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This  Volume  is  for 
REFERENCE  USE  ONLY 


AMERICAN  SOCIETY  of  HEATING 

and  VENTILATING  ENGINEERS 

GUIDE,  1935 


AN  INSTRUMENT  OF  SERVICE  PREPARED  FOR  THE  PROFESSION — 
AND  CONTAINING  REFERENCE  DATA  ON  THE  DESIGN  AND 
SPECIFICATION  OF  HEATING  AND  VENTILATING  SYSTEMS- 
BASED  ON  THE  TRANSACTIONS— THE  INVESTIGATIONS  OF  THE 
RESEARCH  LABORATORY  AND  COOPERATING  INSTITUTIONS — 
AND  THE  PRACTICE  OF  THE  MEMBERS  AND  FRIENDS  OF  THE 

SOCIETY 

TOGETHER  WITH  A 

MANUFACTURERS'    CATALOG    DATA    SECTION    CONTAINING 

ESSENTIAL  AND  RELIABLE  INFORMATION  CONCERNING  MODERN 

EQUIPMENT 

ALSO 
THE  ROLL  OF  MEMBERSHIP  OF  THE  SOCIETY 

WITH 
COMPLETE  INDEXES  TO  TECHNICAL  AND  CATALOG  DATA 

Vol.  13 

l5.oo  PER  COPY 

PUBLISHED  ANNUALLY  BY 

AMERICAN  SOCIETY  of  HEATING  and  VENTILATING 
ENGINEERS 

ji  MADISON  AVENUE       .'.       NEW  YORK,  N.  Y. 


COPYRIGHTS  1935 

BY 
AMERICAN  SOCIETY  OF  HEATING  AND  VENTILATING  ENGINEERS 

AND    BY   IT 

DEDICATED 

To  THE  ADVANCEMENT  OF 
THE  PROFESSION 

AND 

ITS  ALLIED  INDUSTRIES 


TEXT  AND  ILLUSTRATIONS  ARE  FULLY  PRO- 
TECTED BY  COPYRIGHT  AND  NOTHING  THAT 
APPEARS  MAY  BE  REPRINTED  EITHER  WHOLLY 
OR  IN  PART  WITHOUT  SPECIAL  PERMISSION. 


Printed  and  Sound  ly 
THE  HORN-SHAFER  COMPANY 
BALTIMORE     :-:    MARYLAND 


Contents 


Page 

TITLE  PAGE _ i 

CONTENTS Hi 

PREFACE.- , iv 

EDITORIAL  ACKNOWLEDGMENT v 

CODE  OF  ETHICS  FOR  ENGINEERS , vi 

CHAPTER    1.  Fundamentals  of  Heating  and  Air  Conditioning 1 

CHAPTER    2.  Ventilation  and  Air  Conditioning  Standards... 33 

CHAPTER    3.  Industrial  Air  Conditioning 65 

CHAPTER    4.  Natural  Ventilation . 77 

CHAPTER    5.  Heat  Transmission  Coefficients  and  Tables 91 

CHAPTER    6.  Air  Leakage 119 

CHAPTER    7.  Heating  Load 131 

CHAPTER    8.  Cooling  Load 145 

CHAPTER    9.  Central  Air  Conditioning  Systems 155 

CHAPTER  10.  Cooling  Methods 165 

CHAPTER  11.  H modification  and  Dehumidification 183 

CHAPTER  12.  Unit  Air  Conditioners  and  Conditioning  Systems ,....  197 

CHAPTER  13.  Unit  Heaters,  Ventilators,  and  Coolers 219 

CHAPTER  14.  Automatic  Control—. 239 

CHAPTER  15.  Air  Pollution,. 259 

CHAPTER  16.  Air  Cleaning  Devices..... 271 

CHAPTER  17.  Fans  and  Motive  Power. 281 

CHAPTER  18.  Sound  Control ..'. 299 

CHAPTER  19.  Air  Distribution 317 

CHAPTER  20.  Air  Duct  Design 325 

CHAPTER  21.  Industrial  Exhaust  Systems 345 

CHAPTER  22.  Fan  Systems  of  Heating. 359 

CHAPTER  23.  Mechanical  Warm  Air  Furnace  Systems 375 

CHAPTER  24.  Gravity  Warm  Air  Furnace  Systems 389 

CHAPTER  25.  Boilers - 405 

CHAPTER  26.  Chimneys  and  Draft  Calculations 423 

CHAPTER  27.  Fuels  and  Combustion 443 

CHAPTER  28.  Automatic  Fuel  Burning  Equipment 457 

CHAPTER  29.  Fuel  Utilization 479 

CHAPTER  30.  Radiators  and  Gravity  Convectors— 491 

CHAPTER  31.  Steam  Heating  Systems 503 

CHAPTER  32.  Piping  for  Steam  Heating  Systems. 527 

CHAPTER  33.  Hot  Water  Heating  Systems  and  Piping. 559 

CHAPTER  34.  Pipe,  Fittings,  Welding.- 579 

CHAPTER  35.  Water  Supply  Piping. 599 

CHAPTER  36.  Insulation  of  Piping.__ 623 

CHAPTER  37.  District  Heating.. ^39 

CHAPTER  38.  Radiant  Heating ./657 

CHAPTER  39.  Electrical  Heating ^.~  667 

CHAPTER  40.  Test  Methods  and  Instruments... 675 

CHAPTER  41.  Terminology 685 

INDEX  TO  TECHNICAL  DATA 707 

CATALOG  DATA  SECTION - 723 

INDEX  TO  MODERN  EQUIPMENT 947 

INDEX  TO  ADVERTISERS™ 959 

ROLL  OF  MEMBERSHIP.™ 1-57 


PREFACE  TO  THE  13th  EDITION 

THE  ambitious  plans  of  the  Guide  Publication  Committee,  embodying 
several  innovations  to  extend  the  usefulness  of  this  reference  volume, 
have  been  incorporated  in  this  13th  annual  edition  of  THE  AMERICAN 
SOCIETY  OF  HEATING  AND  VENTILATING  ENGINEERS  GUIDE.  The  process 
of  reviewing,  revising  and  reconstructing  the  Technical  Data  Section  and 
then  coordinating  the  complex  subject  matter  of  the  41  chapters  has 
engaged  the  attention  of  over  200  members  so  that  THE  A.S.H.V.E. 
GUIDE  1935  will  appeal  to  an  increasing  number  of  readers  and  give  them 
comprehensive  data  that  are  authoritative  and  practical. 

-Basic  and  fundamental  data  have  been  retained  from  previous  editions 
and  in  those  divisions  where  changes  in  practice  have  been  observed 
modifications  have  been  made  in  the  text  to  bring  the  material  up-to-date. 
The  text  of  THE  GUIDE  1935  now  comprises  two  major  divisions:  the 
subject  matter  of  chapters  and  a  supplementary  section  of  the  problems 
and  answers.  These  problems  and  their  solutions  presented  as  an  appen- 
dix to  each  chapter  represent  the  interpretation  of  the  text  by  a  com- 
petent engineer  whose  analysis  has  been  carefully  reviewed  by  the  Guide 
Publication  Committee.  It  should  be  understood,  however,  that  for 
certain  general  questions,  more  than  one  answer  can  be  made  so  that  the 
addition  of  these  questions  which  represent  problems  in  practice  greatly 
broadens  the  scope  of  THE  GUIDE  and  generally  enhances  its  usefulness. 
As  developments  in  the  manufacturing  field  have  produced  new  appa- 
ratus and  new  applications  of  equipment  for  automatic  heat  and  air 
conditioning  to  improve  comfort,  those  chapters  of  THE  GUIDE  which 
discuss  such  equipment  as  controls,  air  washers,  unit  conditioners,  oil 
burners,  stokers,  etc.,  have  been  reviewed  by  representative  committees 
of  engineers  from  manufacturers'  associations  so  that  the  latest  develop- 
ments in  their  respective  fields  could  be  included. 

The  original  conception  of  THE  GUIDE  outlined  by  its  founders  has 
been  carefully  safeguarded  and  the  aim  of  the  Guide  Publication  Com- 
mittee is  to  have  THE  GUIDE  1935  maintain  its  leadership,  and  continue 
in  its  role,  as  the  recognized  authority  in  the  fields  of  heating,  ventilating 
and  air  conditioning.  Thousands  of  engineers,  architects,  contractors 
and  students  have  come  under  the  influence  of  THE  GUIDE  since  its  first 
appearance  in  1922  and  they  have  found  the  data  authoritative  for  their 
work  in  design,  specification  writing,  installation  or  operation  of  appa- 
ratus and  systems. 

The  Catalog  Data  of  manufacturers  is  nearly  40  per  cent  greater  in 
this  current  edition  indicating  that  THE  GUIDE  is  also  recognized  as  an 
effective  advertising  medium  for  promoting  the  use  of  modern  equipment. 

THE  GUIDE  1935  contains  150  pages  more  than  the  preceding  volume 
and  the  Guide  Publication  Committee  release  this  13th  edition  of 
10,000  copies,  as  a  major  contribution  by  the  Society  toward  the  general 
advancement  of  the  engineering  profession  and  its  allied  industries  in  the 
field  of  heating,  ventilating  and  air  conditioning. 

GUIDE  PUBLICATION  COMMITTEE 
W«  L.  FLEISHER,  Chairman 

JOHN  HOWATT       E.  N.  MCDONNELL 
G,  L.  LARSON       W.  M.  SAWDON 
S.  R.  LEWIS          J.  H.  WALKER 


EDITORIAL  ACKNOWLEDGMENT 


IT  is  with  a  profound  feeling  of  pride  that  the  Guide  Publication 
Committee  acknowledges  the  assistance  and  cooperation  of  the  many 
contributors   to   the  Technical    Data   Section   which  appears  in   THE 
GUIDE  1935. 

A.  J.  NESBITT 

P.  NICHOLLS 

PROF.  L.  S.  O'BANNON 

G,  E.  OLSEN 

G.  H.  OSBORNE 

J.  S.  PARKINSON 

ALBERT  PELLETIER 

E.  C.  RACK 

W.  C.  RANDALL 

P.  L.  REED 

W.  N.  RICH 

PROF.  T.  F.  ROCKWELL 

C.  Z.  ROSECRANS 

PROF.  F.  B.  ROWLEY 

E.  B.  ROYER 
S.  S.  SANFORD 
J.  H.  SCARR 
L.  W.  SCHAD 

W.  G.  SCHLICHTING 

F.  E.  SEDGWICK 
J.  G.  SHODRON 
W.  C.  SMITH 
W.  H.  SMITH 
W.  E.  STARK 

C.  W.  STEWART 

D.  J.  STEWART 

A.  G.  SUTCLIFFE 

D.  L.  TAZE 
L.  A.  TEASDALE 
C.  A.  THINN 
W.  W.  TIMMIS 
C.  L.  TOONDER 
R.  N.  TRANE 
WALTER  TUSCH 
PROF.  G.  L.  TUVE 
W.  M.  WALLACE,  II 
F.  W.  WANDLESS 
PERRY  WEST 
PROF.  C.  P.  YAGLOU 

Special  mention  is  due  the  several  Committee  members  who  acted  as 
division  chairmen  and  who  devoted  long  hours  and  gave  generously  of 
their  knowledge  without  thought  of  compensation  other  than  the  satis- 
faction of  contributing  to  the  advancement  of  the  profession.  The  work 
of  J.  L.  Blackshaw  as  technical  assistant  in  the  detailed  work  of  com- 
pilation was  worthy  of  special  acknowledgment. 


T.  N.  ADLAM 
PROF.  A.  B.  ALGREN 
H.  L.  ALT 
H.  H.  ANGUS 
W.  R.  APPELDOORN 
O.  W.  ARMSPACH 
F.  F.  BAHNSON 
A.  E.  BEALS 
E.  H.  BELING 
PAULINE  BLACKSHAW 
J.  J.  BLOOMFIELD 
BERNARD  BOCK 

C.  A.  BOOTH 

D.  S.  BOYDEN 
J.  J.  BRAUN 
ALBERT  BUENGER 
C.  A.  BULKELEY 

E.  K.  CAMPBELL 
M.  L.  CARR 

R.  E.  CHERNE 

L.  A.  CHERRY 

P.  D.  CLOSE 

J.  F.  S.  COLLINS,  JR. 

R.  P.  COOK 

W.  E.  CRANSTON 

A.  A.  CRIQUI 

J.  M.  DALLAVALLE 

M.  I.  DORFAN 

S.  H.  DOWNS 

T.  F.  DWYER 

PROF.  E.  O.  EASTWOOD 

Louis  ELLIOTT 

J0HN  EVERETTS,  JR. 

PROF,  M.  K.  FAHNESTOCK 

F.  H.  FAUST 
W.  G.  FRANK 
HUGO  FRICKE 
W.  F.  FRIEND 


S.  L.  GOODWIN 

DR.  F.  E.  GIESECKE      * 

W.  A.  GRANT 

DR.  LEONARD  GREENBURG 

HERBERT  HERKIMER 

J.  R.  HERTZLER 

L.  W.  HILDRETH 

DR.  E.  VERNON  HILL 

H.  G.  HILL 

PROF.  J.  D.  HOFFMAN 

J.  H.  HOLTON 

F.  C.  HOUGHTEN 
LLOYD  HOWELL 

PROF.  C.  M.  HUMPHREYS 

H.  F.  HUTZEL 

J.  W.  JAMES 

H.  B,  JOHNS 

R.  E.  JONES 

M.  G.  KERSHAW 

D.  D.  KlMBALL 

DR.  V.  O.  KNUDSEN 

S.  KONZO 

PROF.  A.  P.  KRATZ 

C.  E.  LEWIS 

E.  C.  LLOYD 

G.  W.  MARTIN 

J.  S.  M.  MATHEWSON 
P.  F,  MCDERMOTT 
JOHN  McELGiN 
WILLIAM  McLsisn 
H.  B.  MELLER 
R.  A.  MILLER 
DR.  C.  A.  MILLS 

D.  L.  MILLS 

F.  W.  MORSE 
O.  W.  MOTZ 
H.  C.  MURPHY 
PROF.  D.  W.  NELSON 


s,  Chairman 
GUIDE  PUBLICATION  COMMITTEE 


CODE  of  ETHICS  for  ENGINEERS 

ENGINEERING  work  has  become  an  increasingly  important  factor 
in  the  progress  of  civilization  and  in  the  welfare  of  the  community. 
The  engineering  profession  is  held  responsible  for  the  planning,  construc- 
tion and  operation  of  such  work  and  is  entitled  to  the  position  and 
authority  which  will  enable  it  to  discharge  this  responsibility  and  to 
render  effective  service  to  humanity. 

That  the  dignity  of  their  chosen  profession  may  be  maintained,  it  is 
the  duty  of  all  engineers  to  conduct  themselves  according  to  the  principles 
of  the  following  Code  of  Ethics: 


I — The  engineer  will  carry  on  his  professional  work  in  a  spirit  of  fairness 
to  employees  and  contractors,  fidelity  to  clients  and  employers,  loyalty 
to  his  country  and  devotion  to  high  ideals  of  courtesy  and  personal 
honor. 

2 — He  will  refrain  from  associating  himself  with  or  allowing  the  use  of  his 
name  by  an  enterprise  of  questionable  character. 

3 — He  will  advertise  only  in  a  dignified  manner,  being  careful  to  avoid 
misleading  statements. 

4 — He  will  regard  as  confidential  any  information  obtained  by  him  as  to 
the  business  affairs  and  technical  methods  or  processes  of  a  client  or 
employer. 

5 — He  will  inform  a  client  or  employer  of  any  business  connections,  interests 
or  affiliations  which  might  influence  his  judgment  or  impair  the 
disinterested  quality  of  his  services. 

6 — He  will  refrain  from  using  any  improper  or  questionable  methods  of 
soliciting  professional  work  and  will  decline  to  pay  or  to  accept  com- 
missions for  securing  such  work. 

7 — He  will  accept  compensation,  financial  or  otherwise,  for  a  particular 
service,  from  one  source  only,  except  with  the  full  knowledge  and 
consent  of  all  interested  parties. 

8 — He  will  not  use  unfair  means  to  win  professional  advancement  or  to 
injure  the  chances  of  another  engineer  to  secure  and  hold  employment. 

9 — He  will  cooperate  in  upbuilding  the  engineering  profession  by  exchang- 
ing general  information  and  experience  with  his  fellow  engineers  and 
students  of  engineering  and  also  by  contributing  to  work  of  engineering 
societies,  schools  of  applied  science  and  the  technical  press. 

10 — He  will  interest  himself  in  the  public  welfare  in  behalf  of  which  he  will 
be  ready  to  apply  his  special  knowledge,  skill  and  training  for  the  use 
and  benefit  of  mankind. 


Chapter  1 

FUNDAMENTALS  OF  HEATING  AND 
AIR  CONDITIONING 

Dalton's  Law,  Dry-  and  Wet-Bulb  Temperatures,  Properties  of 
Air,  Humidity,  Relative  Humidity,  Specific  Humidity,  Relation 
of  Dew  Point  to  Relative  Humidity,  Adiabatic  Saturation  of  Air, 
Total  Heat  and  Heat  Content,  Enthalpy,  Psychrometric  Chart, 
Properties  of  Steam,  Properties  of  Water,  Rate  of  Evaporation 

AIR  conditioning  has  for  its  objective  the  supplying  and  maintaining, 
in  a  room  or  other  enclosure,  of  an  atmosphere  having  a  composition, 
temperature,  humidity,  and  motion  which  will  produce  desired  effects 
upon  the  occupants  of  the  room  or  upon  materials  stored  or  handled  in  it. 

Dry  air  is  a  mechanical  mixture  of  gases  composed,  in  percentage  of 
volume,  as  follows1:  nitrogen  78.03,  oxygen  20.99,  argon  0.94,  carbon 
dioxide  0.03,  and  small  amounts  of  hydrogen  and  other  gases. 

Atmospheric  air  at  sea  level  is  given  in  percentage  by  volume  as:  Ns 
77.08,  O2  20.75,  water  vapor  1.2,  A  0.93,  CO2  0.03  and  H2  0.01.  The 
amount  of  water  vapor  varies  greatly  under  different  conditions  and  is 
frequently  one  of  the  most  important  constituents  since  it  affects  bodily 
comfort  and  greatly  affects  all  kinds  of  hygroscopic  materials. 

LAW  OF  PARTIAL  PRESSURES 

A  mixture  of  dry  gases  and  water  vapor,  such  as  atmospheric  air,  obeys 
Dal  ton's  Law  of  Partial  Pressures:  each  gas  or  vapor  in  a  mixture,  at  a 
given  temperature,  contributes  to  the  observed  pressure  the  same  amount 
that  it  would  have  exerted  by  itself  at  the  same  temperature  had  no  other 
gas  or  vapor  been  present.  If  p  —  the  observed  pressure  of  the  mixture 
and  p^  p2,  ps,  etc.  =  the  pressure  of  the  gases  or  vapors  corresponding  to 
the  observed  temperature,  then 

P  =  pi  +  pz  +  p3,  etc.  (1) 

DRY-  AND  WET-BULB  TEMPERATURES 

Air  is  said  to  be  saturated  at  a  given  temperature  when  the  water  vapor 
mixed  with  the  air  is  in  the  dry  saturated  condition  or,  what  is  the  equiv- 
alent, when  the  space  occupied  by  the  mixture  holds  the  maximum  pos- 
sible weight  of  water  vapor  at  that  temperature.  If  the  water  vapor 
mixed  with  the  dry  air  is  superheated,  i.e.,  if  its  temperature  is  above  the 
temperature  of  saturation  for  the  actual  water  vapor  partial  pressure,  the 
air  is  not  saturated. 


^International  Critical  Tobies. 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

The  starting  point  of  most  applications  of  thermodynamic  principles  to 
air-conditioning  problems  is  the  experimental  determination  of  the  dry- 
bulb  and  wet-bulb  temperatures,  and  sometimes  the  barometric  pressure. 

The  dry-bulb  temperature  of  the  air  is  the  temperature  indicated  by  any 
type  of  thermometer  not  affected  by  the  water  vapor  content  or  relative 
humidity  of  the  air.  The  'wet-bulb  temperature  is  determined  by  a  thermo- 
meter with  its  bulb  encased  in  a  fine  mesh  fabric  bag  moistened  with  clean 
water  and  whirled  through  the  arir  until  the  thermometer  assumes  a 
steady  temperature.  This  steady  temperature  is  the  result  of  a  dynamic 
equilibrium  between  the  rate  at  which  heat  is  transferred  from  the  air  to 
the  water  on  the  bulb  and  the  rate  at  which  this  heat  is  utilized  in  evapor- 
ating moisture  from  the  bulb.  The  rate  at  which  heat  is  transferred  from 
the  air  to  the  water  is  substantially  proportional  to  the  wet-bulb  depres- 
sion (t  —  £l),  while  the  rate  of  heat  utilization  in  evaporation  is  propor- 
tional to  the  difference  between  the  saturation  pressure  of  the  water  at 
the  wet-bulb  temperature  and  the  actual  partial  pressure  of  the  water 
vapor  in  the  air  (e]  —  e).  Carriers  equation  for  this  dynamic  equilibrium 
is 


t  -  t1        2800  -  1.3*' 
In  the  form  commonly  used, 


(2a) 
^     J 


2800  -  L3*1 
where 

e  =  actual  partial  pressure  of  water  vapor  in  the  air,  inches  of  mercury. 
e1  -  saturation  pressure  at  wet-bulb  temperature,  inches  of  mercury. 
B  ~  barometric  pressure,  inches  of  mercury. 

t  =  dry-bulb  temperature,  degrees  Fahrenheit. 
/«  ss  wet-bulb  temperature,  degrees  Fahrenheit. 

Formula  2b  may  be  used  to  determine  the  actual  partial  pressure  of  the 
water  vapor  in  a  dry  air-water  vapor  mixture.  Then,  from  Dalton's  Law 
of  Partial  Pressures,  Equation  1,  it  follows  that  the  partial  pressure  of  the 
dry  air  is  (B  —  e). 

If  a  mixture  of  dry  air  and  water  vapor,  initially  unsaturated,  be  cooled 
at  constant  pressure,  the  temperature  at  which  condensation  of  the  water 
vapor  begins  is  called  the  dew-point  temperature.  Clearly  the  dew-point 
is  the  saturation  temperature  corresponding  to  the  actual  partial  pressure, 
e,  of  the  water  vapor  in  the  mixture. 

PROPERTIES  OF  AIR 

Density  is  variously  defined  as  the  mass  per  unit  of  volume,  the  weight 
per  unit  of  volume,  or  the  ratio  of  the  mass,  or  weight,  of  a  given  volume 
of  a  substance  to  the  mass,  or  weight,  of  an  equal  volume  of  some  other 
substance  such  as  water  or  air  under  standard  conditions  of  temperature 
and  pressure.  The  term  specific  gravity  is  more  commonly  used  to  express 
the  latter  relation  but,  when  the  gram  is  taken  as  the  unit  of  mass  and  the 
cubic  centimeter  as  the  unit  of  volume,  density  and  specific  gravity  have 


CHAPTER  I-^-FUNDAMENTALS  OF  HEATING  AND  AIR  CONDITIONING 

% 

the  same  meaning.  The  term  specific  density  is  sometimes  used  to  dis- 
tinguish the  weight  in  pounds  per  cubic  foot;  and  as  here  used,  density  is 
the  weight  in  pounds  of  one  cubic  foot  of  a  substance. 

The  density  of  air  decreases  with  increase  in  temperature  when  under 
constant  pressure.  The  density  of  dry  air  at  70  F  and  under  standard 
atmospheric  pressure  (29.92  in.  of  Hg)  is  approximately  0.075  Ib  (see 
Table  1),  while  that  of  a  mixture  of  air  and  saturated  water  vapor  at  the 
same  temperature  and  barometric  pressure  is  only  about  0.0743  Ib.  In 
the  mixture  the  density  of  the  dry  air  is  0.0731  and  that  of  the  vapor  is 
0.001 15  Ib  (see  Table  2). 

In  order  to  make  comparisons  of  air  volumes  or  velocities  it  is  necessary 
to  reduce  the  observations  to  a  common  pressure  and  temperature  basis. 
The  basic  pressure  is  usually  taken  as  29.92  in.  of  Hg,  but  no  basic  tem- 
perature is  universally  recognized.  Common  temperatures  for  this 
purpose  are  32  F,  60  F,  68  F,  and  70  F.  Since  70  F  is  the  most  commonly 
specified  temperature  to  which  rooms  for  human  occupancy  must  be 
heated,  it  is  usually  understood,  when  no  other  temperature  is  specified, 
that  70  F  is  the  basic  temperature  for  measuring  the  volume  or  the 
velocity  of  air  in  heating  and  ventilating  work. 

The  specific  volume  of  air  is  the  volume  in  cubic  feet  occupied  by  one 
pound  of  the  air.  Under  constant  pressure  the  specific  volume  varies 
inversely  as  the  density  and  directly  as  the  absolute  temperature. 

The  specific  heat  of  air  is  the  number  of  Btu  required  to  raise  the 
temperature  of  1  Ib  of  air  1  F.  The  specific  heat  at  constant  pressure, 
Cp,  and  that  at  constant  volume,  Cv,  are  different.  The  specific  heat 
at  constant  pressure  is  commonly  used  and  it  varies,  under  a  pressure 
of  one  atmosphere,  from  a  minimum  at  about  32  F  from  which  it  increases 
with  either  increase  or  decrease  of  temperature.  The  value  0.24  is  suf- 
ficiently accurate  for  use  at  ordinary  temperatures,  but  the  values  range1 
from  0.2399  at  32  F  to  0.2404  at  212  F,  0.2413  at  392  F,  0.243  at  - 108  F, 
and  0.252  at  -301  F. 

The  mean  specific  heat  of  water  vapor  at  constant  pressure  is  taken  as 
0.45  for  all  general  engineering  computations. 

Table  3  is  intended  to  aid  in  determining  the  density  of  moist  air, 
taking  into  account  its  temperature,  pressure,  and  moisture  content. 

Example  1.  To  show  the  use  of  Table  3:  Given  air  at  83  F  dry-bulb  and  68  F  wet- 
bulb  (or  a  depression  of  15  deg)  with  a  barometric  pressure  of  29.40  in.  of  mercury. 
What  will  be  the  weight  of  this  air  in  pounds  per  cubic  foot? 

Solution.  From  Table  3  the  weight  of  saturated  air  at  80  F  and  29.00  in.  barometer  is 
found  to  be  0.07034  Ib  per  cubic  foot.  There  is  a  decrease  of  0.00015  Ib  per  degree  dry- 
bulb  temperature  above  80  F.  There  is  an  increase  of  0.00025  Ib  for  each  0.1  in.  above 
29.00  in.  From  the  last  column  of  Table  3  it  is  found  that  there  is  an  increase  of  approxi- 
mately 0.000035  Ib  per  degree  wet-bulb  depression  when  the  dry-bulb  is  83  F.  Tabu- 
lating the  items: 

0.07034  =  weight  of  saturated  air  at  80  F  and  29.00  bar. 
-  0.00045  =  decrement  for  3  deg  dry-bulb,  3  X  0.00015. 
+  0.00100  =  increment  for  0.4  in.  bar.,  4  X  0.00025. 
-f  0.00053  =  increment  for  15  deg  wet-bulb  depression,  15  X  0.000035. 

0.07142  —  weight  in  pounds  per  cubic  foot  of  air  at  83  F  dry-bulb,  68  F  wet-bulb, 
29.40  in.  bar. 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


TABLE  1.    PROPERTIES  OF  DRY  Ama 

Barometric  Pressure  29.921  In. 


TEMPERATURE 
DBS  P 

WEIGHT  PER  Cu  FT 
POUNDS 

PER  CENT  OP  VOLUME 

AT70F 

BTXT  ABSORBED  BY 
ONE  Cu  FT  DRY  AIR 
PER  DEG  F 

Cu  FT  DRY  Am 
WARMED  ONE  DEGREE 
PER  BTU 

0 

0.08636 

0.8680 

0.02080 

48.08 

10 

0.08453 

0.8867 

0.02039 

49.05 

20 

0.08276 

0.9057 

0.01998 

50.05 

30 

0.08107 

0.9246 

0.01957 

51.10 

40 

0.07945 

0.9434 

0.01919 

52.11 

50 

0.07788 

0.9624 

0.01881 

53.17 

60 

0.07640 

0.9811 

0.01846 

54.18 

70 

0.07495 

1.0000 

0.01812 

55.19 

80 

0.07356 

1.0190 

0.01779 

56.21 

90 

0.07222 

1.0380 

0.01747 

57.25 

100 

0.07093 

1.0570 

0.01716 

58.28 

110 

0.06968 

1.0756 

0.01687 

59.28 

120 

0.06848 

1.0945 

0.01659 

60.28 

130 

0.06732 

1.1133 

0.01631 

61.32 

140 

0.06620 

1.1320 

0.01605 

62.31 

150 

0.06510 

1.1512 

0.01578 

63.37 

160 

0.06406 

1.1700 

0.01554 

64.35 

180 

0.06205 

1.2080 

0.01506 

66.40 

200 

0.06018 

1.2455 

0.01462 

68.41 

220 

0.05840 

1.2833 

0.01419 

70.48 

240 

0.05673 

1.3212 

0.01380 

72.46 

260 

0.05516 

1.3590 

0.01343 

74.46 

280 

0.05367 

1.3967 

0.01308 

76.46 

300 

0.05225 

1.4345 

0.01274 

78.50 

350 

0.04903 

1.5288 

0.01197 

83.55 

400 

0.04618 

1.6230 

0.01130 

88.50 

450 

0.04368 

1.7177 

0.01070 

93.46 

500 

0.04138 

1.8113 

0.01018 

98.24 

550 

0.03932 

1.9060 

0.00967 

103.42 

600 

0.03746 

2.0010 

0.00923 

108.35 

700 

0.03423 

2.1900 

0.00847 

11$.  07 

800 

0.03151 

2.3785 

0.00782 

127.88 

900 

0.02920 

2.5670 

0.00728 

137.37 

1000 

0.02720 

2.7560 

0.00680 

147.07 

•From  Fan  Engineering* 


CHAPTER  1  —  FUNDAMENTALS  OF  HEATING  AND  AIR  CONDITIONING 

It  is  usual  to  assume  that  dry  air,  moist  air,  and  the  water  vapor  in  the 
air  follow  the  laws  of  perfect  gases.  This  assumption  while  not  absolutely 
true,  especially  with  saturated  vapor  at  temperatures  much  above  140  F, 


TABLE  2.    PROPERTIES  OF  SATURATED  Araa 

Weights  of  Air,  Vapor  of  Water,  and  Saturated  Mixture  of  Air  and  Vapor  at  29.921  Inches  of  Mercury 


TEMP. 
DEG.F 

WEIGHT  IN  A  CTTBIC  FOOT  OF  MITTUBE 

BTU  ABSORBED  BT 
ONE  CUBIC  FOOT 
SAT.  Am  PER 
DEGF 

CUBIC  FEET  SAT. 
Am  WABMED  ONE 
DEGREE  PER 
BTU 

SPECIFIC 
HEAT  BTU 
PER  POUND 

OlMlXTUKI 

WEIGHT  OP 
DRY  Am 
POUNDS 

WEIGHT  OP 
VAPOE 
POUNDS 

TOTAL  WEIGHT  OP 
THE  MDCTURE 
POUNDS 

0 

0.08625 

0.000068 

0.08632 

0.02083 

48.02 

0.2413 

10 

0.08433 

0.000110 

0.08444 

0.02039 

49.05 

0.2415 

20 

0.08246 

0.000176 

0.08264 

0.01998 

50.07 

0.2418 

30 

0.08062 

0.000277 

0.08090 

0.01958 

51.07 

0.2420 

40 

0.07878 

0.000409 

0.07919 

0.01921 

52.06 

0.2426 

50 

0.07694 

0.000587 

0.07753 

0.01885 

53.05 

0.2431 

60 

0.07506 

0.000828 

0.07589 

0.01851 

54.02 

0.2439 

70 

0.07310 

0.001151 

0.07425 

0.01819 

54.97 

0.2450 

80 

0.07103 

0.001578 

0.07261 

0.01790 

55.87 

0.2465 

90 

0.06879 

0.002134 

0.07092 

0.01762 

56.76 

0.2485 

100 

0.06635 

0.002850 

0.06920 

0.01736 

57.59 

0.2509 

110 

0.06364 

0.003762 

0.06740 

0.01714 

58.35 

0.2543 

120 

0.06060 

0.004914 

0.06551 

0.01695 

59.00 

0.2587 

130 

0.05715 

0.006351 

0.06350 

0.01679 

59.56 

0.2644 

140 

0.05319 

0.008120 

0.06131 

0.01668 

59.96 

0.2721 

150 

0.04864 

0.010295 

0.05894 

0.01662 

60.17 

0.2820 

160 

0.04340 

0.012936 

0.05634 

0.01662 

60.17 

0,2950 

170 

0.03734 

0.016108 

0.05345 

0.01668 

59.96 

0.3121 

ISO 

0.03035 

0.019896 

0.05025 

0.01684 

59.38 

0.3351 

190 

0.02228 

0.024400 

0.04668 

0.01710 

58.49 

0.3663 

200 

0.01300 

0.029715 

0.04272 

0.01749 

57.18 

0.4094 

210 

0.00230 

0.035938 

0.03824 

0.01802 

55.50 

0.4712 

212 

0.00000 

0.037307 

0.03731 

0.01815 

55.10 

0.4865 

aFroip.  Fan  Engineering. 

is  sufficiently  accurate  for  practical  purposes  and  it  greatly  simplifies 
computations. 

Boyle's  Law  refers  to  the  relation  between  the  pressure  and  volume  of  a 
gas,  and  may  be  stated  as  follows :  With  temperature  constant,  the  volume  of 
a  given  weight  of  gas  varies  inversely  as  its  absolute  pressure.  Hence,  if 
PI  and  P2  represent  the  initial  and  final  absolute  pressures,  and  V\  and 
F2  represent  corresponding  volumes  of  the  same  mass,  say  one  pound  of 

V       P 
gas,  then  •=?  =  --,  or  PI  FI  =  P2  F2,  but  since  PI  FI  for  any  given  case  is 

a  definite  constant  quantity,  It  follows  that  the  product  of  the  absolute 

5 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


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CHAPTER  1 — FUNDAMENTALS  OF  HEATING  AND  AIR  CONDITIONING 

pressure  and  volume  of  a  gas  is  a  constant,  or  PV  =  C,  when  T  is  kept 
constant.  Any  change  in  the  pressure  and  volume  of  a  gas  at  constant 
temperature  is  called  an  isothermal  change. 

Charles1  Law  refers  to  the  relation  among  pressure,  volume,  and  tem- 
perature of  a  gas  and  may  be  stated  as  follows:  The  volume  of  a  given 
weight  of  gas  varies  directly  as  the  absolute  temperature  at  constant  pressure, 
and  the  pressure  varies  directly  as  the  absolute  temperature  at  constant 
volume.  Hence,  when  heat  is  added  at  constant  volume,  Fc,  the  resulting 

~P          T 

equation  is  ~  =  •—,  or,  for  the  same  temperature  range  at  constant  pres- 
-t  i      l\ 

sure,  PC,  the  relation  is  ~  =  •—. 

In  general,  for  any  weight  of  gas,  W,  since  volume  is  proportional  to 
weight,  the  relation  among  P,  V,  and  T  is 

PV  =  WRT  (3) 

where 

P  —  the  absolute  pressure  of  the  gas,  pounds  per  square  foot. 
V  =  the  volume  of  the  weight  W,  cubic  feet. 
W  —  the  weight  of  the  gas,  pounds. 
R  =  a  constant  depending  on  the  nature  of  the  gas.    The  average  value  of  R  for  air 

is  53.34. 
T  =  the  absolute  temperature,  degrees  Fahrenheit. 

This  is  the  characteristic  equation  for  a  perfect  gas,  and  while  no  gases 
are  perfect  in  this  sense,  they  conform  so  nearly  that  Equation  3  will 
apply  to  most  engineering  computations. 

HUMIDITY 

Humidity  is  the  water  vapor  mixed  with  dry  air  in  the  atmosphere. 
Absolute  humidity  has  a  multiplicity  of  meanings,  but  usually  the  term 
refers  to  the  weight  of  water  vapor  per  unit  volume  of  space  occupied, 
expressed  in  grains  or  pounds  per  cubic  foot.  With  this  meaning,  absolute 
humidity  is  nothing  but  the  actual  density  of  the  water  vapor  in  the 
mixture  and  might  better  be  so  called.  A  study  of  Keenan's  Steam 
Tables2  indicates  that  water  vapor,  either  saturated  or  super-heated,  at 
partial  pressures  lower  than  4  in.  of  mercury  may  be  treated  as  a  gas  with 
a  gas  constant  R  of  1.21  in  the  characteristic  equation  of  the  gas  pV  = 
wR  (t  +  460).  Within  such  limits,  the  density  (8)  of  water  vapor  is 

(pounds  per  cubic  foot)  (4a) 


1.21  (t  +  460) 

5785  e   (grains  per  cubic  foot)  (4b) 


t  +  460 
where 

e  =  actual  partial  pressure  of  vapor,  inches  of  mercury* 
t  =  dry-bulb  temperature,  degrees  Fahrenheit. 


•Published  by  American  Society  of  Mechanical  Engineers,  see  abstract  in  Table  7. 

7 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

Specific  Humidity 

It  simplifies  many  problems  which  deal  with  mixtures  of  dry  air  and 
water  vapor  to  express  the  weight  or  the  mass  of  the  vapor  in  terms  of  the 
weight  or  the  mass  of  dry  air.  If  the  weight  of  the  water  vapor  in  a 
mixture  be  divided  by  the  weight  of  the  dry  air,  and  the  weight  of  dry  air 
be  made  unity,  we  have  an  expression  of  the  weight  of  water  vapor  carried 
by  a  unit  weight  of  dry  air.  This  relation  has  no  generally  accepted  name. 
It  has  been  variously  called:  mixing  ratio,  proportionate  humidity,  mass 
or  density  ratio,  absolute  humidity,  and  specific  humidity.  Of  all  these 
terms  specific  humidity  is  the  most  suggestive  of  the  meaning  which  it  is 
desired  to  express  and  it  has  found  considerable  use  in  this  sense  even 
though  it  is  defined  in  International  Critical  Tables  as  the  ratio  of  the 
mass  of  vapor  to  the  total  mass.  It  will  be  understood  here  that  specific 
humidity  refers  to  the  weight  of  water  vapor  in  pounds  carried  by  one 
pound  of  dry  air. 

The  gas  constant  for  dry  air,  when  the  partial  pressure  of  the  air  is 
expressed  in  inches  of  Hg,  is  0.753;  so  that  the  specific  humidity,  if 
represented  by  IF,  is 

w/  e  -  B~e 

W  = 


1.21  (/  H-  460)    '    0.753  (/  +  460) 
=  0.622  (~-\  (pounds)  (5a) 


=  4354  (  jl~\  (grains)  (5b) 

where 

e  =  actual  partial  pressure  of  vapor,  inches  of  mercury. 
B  =  total  pressure  of  mixture  (barometric  pressure),  inches  of  mercury. 

Relative  Humidity 

Relative  humidity  ($)  is  either  the  ratio  of  the  actual  partial  pressure, 
e,  of  the  water  vapor  in  the  air  to  the  saturation  pressure,  et,  at  the  dry- 
bulb  temperature,  or  the  ratio  of  the  actual  density,  8,  of  the  vapor  to 
the  density  of  saturated  vapor,  8t,  at  the  dry-bulb  temperature.  That  is: 

*-i  =  i  (6) 

The  relative  humidity  of  a  given  mixture  at  af given  temperature  is  not 
the  same  as  the  specific  humidity,  Wt  of  the  mixture  divided  by  the 
specific  humidity,  Wt,  of  saturated  vapor  at  the  same  temperature,  for 
from  Equations  5a  and  6 


0.622         -     -  (7) 


<  _  .  - 

Wt  \P  —  3>  et/  \  B-et  )       B  — 

The  specific  humidity  of  an  unsaturated  air-vapor  mixture  cannot, 
therefore,  be  accurately  found  by  multiplying  the  specific  humidity  of 
saturated  vapor  by  its  relative  humidity;  although  the  error  is  usually 
small  especially  when  the,  relative  humidity  is  high. 

With  a  relative  humidity  of  100  per  cent,  the  dry-bulb,  wet-bulb,  and 

8 


CHAPTER  1 — FUNDAMENTALS  OF  HEATING  AND  AIR  CONDITIONING 


dew-point  temperatures  are  equal.  With  a  relative  humidity  less  than 
100  per  cent,  the  dry-bulb  exceeds  the  wet-bulb,  and  the  wet-bulb  exceeds 
the  dew-point  temperature. 

RELATION  OF  DEW  POINT  TO  RELATIVE,  HUMIDITY 

A  peculiar  relationship  exists  between  the  dew  point  and  the  relative 
humidity  and  this  is  found  most  useful  in  air  conditioning  work.  This 
relationship  is,  that  for  a  fixed  relative  humidity  there  is  substantially  a 
constant  difference  between  the  dew  point  and  the  dry-bulb  temperature 
over  a  considerable  temperature  range.  Table  4,  giving  the  dry-bulb  and 
dew-point  temperatures  and  the  dew-point  differentials  for  50  per  cent 
relative  humidity,  illustrates  this  relationship  clearly. 


TABLE  4. 


DRY-BULB  AND   DEW-POINT  TEMPERATURES  FOR 
50  PER  CENT  RELATIVE  HUMIDITY 


Dry-bulb  temperature  

65.0 

70.0 

75.0 

80.0 

85.0 

90.0 

Dew-point  temperature 

45.8 

50.5 

55.25 

59.75 

64.25 

68.75 

Difference  between  dew-point  and  dry- 
bulb  temperature  

19.2 

19.5 

19.75 

20.25 

20.75 

21.25 

It  will  be  seen  from  an  inspection  of  this  table  that  the  difference 
between  the  dew-point  temperature  and  the  room  temperature  is  approxi- 
mately 20  deg  throughout  this  range  of  dry-bulb  temperatures  or,  to 
be  more  exact,  the  differential  increases  only  10  per  cent  for  a  range  of 
practically  25  deg. 

This  principle  holds  true  for  other  humidities  and  is  due  to  the  fact 
that  the  pressure  of  the  water  vapor  practically  doubles  for  .every  20  deg 
through  this  range* 

The  approximate  relative  humidity  for  any  difference  between  dew- 
point  and  dry-bulb  temperature  may  be  expressed  in  per  cent  as: 


100 


(8) 


where 


dew-point  temperature. 


This  principle  is  very  useful  in  determining  the  available  cooling  effect 
obtainable  with  saturated  air  when  a  desired  relative  humidity  is  to  }>e 
maintained  in  a  room,  even  though  there  may  be  a  wide  variation  in  room 
temperature.  This  problem  is  one  which  applies  to  certain  industrial  con- 
ditions, such  as  those  in  cotton  mills  and  tobacco  factories,  where  re- 
latively high  humidities  are  carried  and  where  one  of  the  principal  prob- 
lems is  to  remove  the  heat  generated  by  the  machinery.  It  also  permits 
the  use  of  a  differential  thermostat,  responsive  to  both  the  room  tempera- 
ture and  the  dew-point  temperature,  to  control  the  relative  humidity 
in  the  room. 

Table  5  gives,  for  different  temperatures,  the  density  of  saturated  vapor, 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


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10 


CHAPTER  1 — FUNDAMENTALS  OF  HEATING  AND  AIR  CONDITIONING 


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11 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


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12 


CHAPTER  1 — FUNDAMENTALS  OF  HEATING  AND  AIR  CONDITIONING 


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S2S I 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

St,  the  weight  of  saturated  vapor  mixed  with  1  Ib  of  dry  air,  Wt,  (at  a 
relative  humidity  of  100  per  cent  and  a  barometric  pressure,  B,  of  29.92  in. 
of  mercury) ,  the  specific  volume  of  dry  air,  and  the  volume  of  an  air- vapor 
mixture  containing  1  Ib  of  dry  air  (at  a  relative  humidity  of  100  per  cent 
and  a  pressure  of  29.92  in.  of  mercury).  The  preceding  equations  or  the 
data  from  Table  5  may  be  conveniently  used  in  solving  the  following 
typical  problems :  (See  Table  6  for  temperatures  below  OF.) 

Example  2.  Humidifying  and  Heating.  Air  is  to  be  maintained  at  70  F  with  a  relative 
humidity  of  40  per  cent  (3?  —  0.4)  when  the  outside  air  is  at  0  F  and  70  per  cent 
relative  humidity  (<£  ==  0.7)  and  a  barometric  pressure,  B,  of  29.92  in.  of  mercury.  Find 
the  weight  of  water  vapor  added  to  each  pound  of  dry  air  and  the  dew-point  temperature 
of  the  humidified  air. 

Solution.    From  Equation  5a  and  Table  5, 

0.622  X_Q°  =  °-000547  lb  Per  P°und  of  dT  air- 

°*°0618  lb  per  P°Und  °f  dry  air' 

The  water  vapor  added  per  pound  of  dry  air  must  be  (Wz  -  Wi)  or  0.005633  lb.  By 
inspection  of  Table  5,  Wt  =  0.00618  at  44.5  F,  so  this  is  the  dew-point  temperature  of 
the  humidified  air. 

An  approximation  of  the  same  result  from  Table  5  is 

Wi  =  0.7  X  0.000781  »  0.000547  lb  per  pound  of  dry  air. 
W2  =  0.4  X  0.01578  =  0.006312  lb  per  pound  of  dry  air. 

The  water  vapor  added  per  pound  of  dry  air  is  approximately  0.005765  lb  and  the 
dew-point  temperature  is  approximately  45  F.  The  degree  of  approximation  is  evident. 

Example  #.  Dehumidifying  and  Cooling.  Air  with  a  dry-bulb  temperature  of  84  F, 
a  wet-bulb  of  70  F,  or  a  relative  humidity  of  50  per  cent  (<3>  =  0.5),  and  a  barometric 
pressure,  5,  of  29.92  in.  of  mercury  is  to  be  cooled  to  54  F.  Find  the  dew-point  tem- 
perature of  the  entering  air  and  the  weight  of  vapor  condensed  per  pound  of  dry  air. 

Solution.    From  Equation  5a  and  Table  5, 

Wi  =  0.622  (29  ^-^Q1 587)  =  °-01245  lb  Per  P°und  of  ^  ain 

w,  =  0.622  (2992^042)  "  °-00887  lb  Per  p°und  of  dry air- 

Since  Wi  =  Wt  when  /  =  63.3  F,  this  is  the  dew-point  temperature  of  the  entering  air. 
The  weight  of  vapor  condensed  is  (W\  —  Wz)  or  0.00358  lb  per  pound  of  dry  air. 

An  approximate  result  is 

Wi  =  0.5  X  0.02547  =  0.01274  lb  per  pound  of  dry  air. 

Wi  =  1  X  0.00887  =  0.00887  lb  per  pound  of  dry  air,  since  the  exit  air  is  saturated. 

Since  Wi  =  Wt  at  t  -  64  F,  this  is  the  dew-point  temperature  of  the  entering  air. 
The^weight  of  vapor  condensed  is  0.00387  lb  per  pound  of  dry  air.  The  degree  of  approxi- 
mation is  again  evident. 

ADIABATIC  SATURATION  OF  AIR 

The  process  of  adiabatic  saturation  of  air  is  of  considerable  importance 
in  air-conditioning.  Suppose  that  1  lb  of  dry  air,  initially  unsaturated  but 
carrying  W  lb  of  water  vapor  with  a  dry-bulb  temperature,  t,  and  a  wet- 

14 


CHAPTER  1 — FUNDAMENTALS  OF  HEATING  AND  AIR  CONDITIONING 

bulb  temperature,  f,  be  made  to  pass  through  a  tunnel  containing  an 
exposed  water  surface.  Further  assume  the  tunnel  to  be  completely  in- 
sulated, thermally,  so  that  the  only  heat  transfer  possible  is  that  between 
the  air  and  water.  As  the  air  passes  over  the  water  surface,  it  will  gradu- 
ally pick  up  water  vapor  and  will  approach  saturation  at  the  initial  wet- 
bulb  temperature  of  the  air,  if  the  water  be  supplied  at  this  wet-bulb  tem- 
perature. During  the  process  of  adiabatic  saturation,  then,  the  dry-bulb 
temperature  of  the  air  drops  to  the  wet-bulb  temperature  as  a  limit,  the 
wet-bulb  temperature  remains  substantially  constant,  and  the  weight  of 
water  vapor  associated  with  each  pound  of  dry  air  increases  to  Wv,  as  a 
limit,  where  Wv  is  the  weight  of  saturated  vapor  per  pound  of  dry  air  for 
saturation  at  the  wet-bulb  temperature. 

Example  4-  If  air  with  a  dry-bulb  of  85  F  and  a  wet-bulb  of  70  F  be  saturated  adia- 
batically  by  spraying  with  recirculated  water,  what  will  be  the  final  temperature  and  the 
vapor  content  of  the  air? 

Solution.  The  final  temperature  will  be  equal  to  the  initial  wet-bulb  temperature  or 
70  F,  and  since  the  air  is  saturated  at  this  temperature,  from  Table  5,  W  =  0.01578  Ib 
per  pound  of  dry  air. 

In  the  adiabatic  saturation  process,  since  the  heat  given  up  by  the  dry 
air  and  associated  vapor  in  cooling  to  the  wet-bulb  temperature  is  utilized 
in  evaporation  of  water  at  the  wet-bulb  temperature,  W.  H.  Carrier  has 
pointed  out3  that  the  equation  for  the  process  of  adiabatic  saturation,  and 
hence  for  a  process  of  constant  wet-bulb  temperature,  is: 

fc'fg  (Wti  -  W)  -  cPa  (t  -  *')  +  c^W  (t  ~  *')  (9a) 

and  using  cPa  =  0.24  and  cPs  =  0.45 

#fc  (Wv  -  W)  =  (0.24  -f  0.4517)  (t  -  f)  (9b) 

where 

h*fs  —  latent  heat  of  vaporization  at  t1,  Btu  per  pound. 

(Wt*  —  W)  =  increase  in  vapor  associated  with  1  Ib  of  dry  air  when  it  is  saturated 
adiabatically  from  an  initial  dry-bulb  temperature,  /,  and  an  initial  vapor  content,  W, 
pounds. 

Knowing  any  two  of  the  three  primary  variables,  /,  t',  or  W,  the  third 
may  be  found  from  this  equation  for  any  process  of  adiabatic  saturation. 

TOTAL  HEAT  AND  HEAT  CONTENT 

The  total  heat  of  a  mixture  of  dry  air  and  water  vapor  was  originally 
defined  by  W.  H.  Carrier  as 

S  =  <;Pa  (t  -  0)  -f  W  [fc'fg  +  cPs  (t  -  *')]  (10) 

where 

2  =  total  heat  of  the  mixture,  Btu  per  pound  of  dry  air. 
Cp^  =  mean  specific  heat  at  constant  pressure  of  dry  air. 
Cpg  =s  mean  specific  heat  at  constant  pressure  of  water  vapor. 
t  =  dry-bulb  temperature,  degrees  Fahrenheit. 

#  =  wet-bulb  temperature,  degrees  Fahrenheit. 


*A.SM.E.  Transactions,  Vol.  33,  1911,  p.  1005. 

15 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


TABLE  6.    PROPERTIES  OF  SATURATED  WATER  VAPOR  AT  LOW/TEMPERATURES** 
Barometer,  29.92  Inches  of.  Mercury 


WEIGHT  OP 

WEIGHT  OP 

TEMPERA- 

VAPOR 

PRESSURE 

SATURATED 
VAPOR 

BTU  PER  LB 
OF  VAPOR 

TEMPERA- 

VAPOR 

PBESSURB 

SATURATED 
VAPOR 

BTU  PER  LB 
OF  VAPOR 

TUBE 

IN 

PER  LB 

(32  F 

TURE 

IN. 

PBRLfi 

(32  F 

F 

Hex  10-6 

DRY  AIR 

DATUM) 

F 

HoX  10-6 

DRY  AIR 

DATUM) 

X1QJ 

X  10-6 

-130 

0.276 

0.005738 

1000.7 

-85 

15.87 

0.3299 

1021.0 

-129 

.306 

.006362 

1001.2 

-84 

17.20 

-      .3576 

•1021.4 

-128 

.338 

.007027 

1001.6 

-83 

18.58 

,.3863 

1021.9 

-127 

.373 

.007755 

1002.1 

-82 

20.10 

.4179 

1022.3 

-126 

.411 

.008545 

1002.5 

-81 

21.72 

.4516 

1022.8 

-125 

.455 

.009459 

1003.0 

-80- 

23.47 

.4879 

1023.2 

-124 

.499 

.01037 

1003.4 

-79 

25.34 

.5268 

1023.7 

-123 

.542 

.01127 

1003.9 

-78 

27.29 

.5674 

1024.1 

-122 

-.604 

.01256 

1004.3 

-77 

29.52 

-  .6137 

1024.6 

-121 

.669 

.01391 

1004.8 

-76 

31.81 

.6613 

1025.0 

-120 

.735 

.01528 

1005.2 

-75 

34.37 

.7146 

1025.5 

-119 

.805 

.01674 

1005.7 

-74 

37.01 

.7694 

1025.9 

-118 

.892 

.01854 

1006.1 

-73 

39.96 

.8308 

1026.4 

-117 

.989 

.02056 

1006.6 

-72 

43.04 

.8948 

1026.8 

-116 

1.098 

.02283 

1007.0 

-71 

46.33 

.9632 

1027.3 

-115 

'1.208 

.02511 

1007.5 

-70 

49.87 

1.037 

1027.7 

-114 

1.317 

.02738 

1007.9 

-69 

53.59 

1.114 

1028.2 

-113 

1.444 

.03002 

1008.4 

-68 

57.65 

1.199 

1028.6 

-112 

1.575 

.03274 

1008.8 

-67 

61.81 

1.285 

1029.1 

-111 

1.728 

.03593 

1009.3 

-66 

66.41 

1.381 

1029.5 

-110 

1.889 

.03927 

1009.7 

-65 

71.17 

1.480 

1030.0 

-109 

2.087 

.04339 

1010.2 

-64 

76.64 

1.593 

1030.4 

-108 

2.292 

.04765 

1010.6 

-63 

82.28 

1.711 

1030.9 

-107 

2.511 

.05220 

1011.1 

-62 

88.19 

•1.833 

1031.3 

-106 

2.742 

.05701 

1011.5 

-61 

94.62 

1.967 

1031.8 

-105 

2.983 

.06202 

1012.0 

-60 

101.4 

2.108 

1032.2  * 

-104 

3.258 

.06773 

1012.4 

-59 

108.8 

2.262 

1032.7 

-103 

3.543 

.07366 

1012.9 

-58 

116.3 

2.418 

1033.1 

-102 

'3.872 

.08050 

1013.3 

-57 

124.8 

2.595 

1033.6 

-101 

4.213 

.08759 

1013.8 

-56 

133.4 

2.773 

1034.0 

-100 

4.607 

.09578 

1014.2 

-55 

143.0 

2.973 

1034.5 

-99 

5.018 

.1043 

1014.7 

-54 

153.0 

3.181 

1034.9 

-98 

5.455 

.1134 

1015.1 

-53 

163.5 

3.399 

1035.4 

-97 

5.946 

.1236 

1015.6 

-52 

174.9 

3.636 

1035.8 

-96 

6.470 

.1345 

1016.0 

-51 

187.0 

3.888 

1036.3 

-95 

7.047 

.1465 

1016.5 

-50 

199.9 

4.156 

1036.7 

-94 

7.638 

.1588 

1016.9 

-49 

213.0 

4.428 

1037.2 

-93 

8.316 

.1729 

1017.4 

-48 

227.9 

4.738 

1037.6 

-92 

9.017 

.1875 

1017.8 

-47 

243.1 

5.054 

1038.1 

'  -91 

9.806 

.2039 

1018.3 

-46 

259.5 

5.395 

1038.5 

-90 

10.64 

.2212 

1018.7 

-45 

276.7 

5.753 

1039.0 

-89 

11.53 

.2397 

1019.2 

-44 

295.0 

6.133 

1039.4 

-88 

12.51 

.2601 

1019.6 

-43 

314.7 

6.543 

1039.9 

-87 

13.53 

.2813 

1020.1 

-42 

335.3 

6.971 

1040.3 

-86 

14.69 

.3054 

1020.5 

-41 

357.6 

7.435 

1040.8 

"  "Vapor  pressures  converted  from  International  Critical  Tables. 

16 


CHAPTER  1— FUNDAMENTALS  OF  HEATING  AND  AIR  CONDITIONING 


TABLE  6.  PROPERTIES  OF  SATURATED  WATER  VAPOR  AT  Low  TEMPERATURES**   (Con'd.) 
Barometer,  29.92  Inches  of  Mercury 


WEIGHT  OF 

WEIGHT  OF 

TEMPEBA- 

TTJRE 

F 

VAPOB 

PEESSTJHE 
IN. 
Ho  X  10-5 

SATURATED 
VAPOR 

PERL-B 

DRT  Am 

Bru  PEE  LB 
OP  VAPOR 
(32  F 
DATUM) 

TEMPERA- 
TURE 
F 

VAPOR 
PRESSURE 
IN. 
EG  X  10-5 

SATURATED 
VAPOR 

FERliB 

DRY  AIR 

BTU  PER  LB 
OF  VAPOR 
(32  F 
DATUM) 

X  10-5 

X  10-s 

-40 

380.3 

7.907 

1041.2 

-20 

1262.0 

26.25 

1050.2 

-39 

405.5 

8.431 

1041.7 

-19 

1337. 

27.81 

1050.7 

-38 

431.2 

8.965 

1042.1 

-18 

1416. 

29.45 

1051.1 

-37 

459.2 

9.548 

1042.6 

-17 

1496. 

31.12 

1051.6 

-36 

488.4 

10.16 

1043.0 

-16 

1584. 

32.95 

1052.0 

-35 

519.5 

10.80 

1043.5 

-15 

1675. 

34.84 

1052.5 

-34 

552.4 

11.49 

1043.9 

-14 

1772. 

36.86 

1052.9 

-33 

586.5 

12.20 

1044.4 

-13 

1874. 

38.98 

1053.4 

-32 

623.7 

12.97 

1044.8 

-12 

1980. 

41.19  ' 

1053.8 

-31 

661.8 

13.76 

1045.3 

-11 

2093. 

43.54 

1054.3 

-30 

701.0 

14.58 

1045.7 

-10 

2210. 

45.98 

1054.7 

-29 

742.2 

15.43 

1046.2 

-9 

2335. 

48.58 

1055.2 

-28 

791.2 

16,45 

1046.6 

-8 

2463. 

51.25 

1055.6 

-27 

841.0 

17.49 

1047.1 

-7 

2502. 

52.06 

1056.1 

-26 

892.1 

18.55 

1047.5 

-6 

2745. 

57.12 

1056.5 

-25 

946.4 

19.68 

1048.0 

-5 

2898. 

60.30 

1057.0 

-24 

1003. 

20.86 

1048.4 

-4 

3055. 

63.57 

1057.4 

-23 

1064. 

22.13 

1048.9 

-3 

3222. 

67.05 

1057.9 

-22 

1126. 

23.42 

1049.3 

-2 

3397. 

70.69 

1058.3 

-21 

1192. 

24.79 

1049.8 

-1 

3580. 

74.50 

1058.8 

0 

3773. 

78.52 

1059.2 

a  Vapor  pressures  converted  from  International  Critical  Tables, 


W  =  weight  of  water  vapor  mixed  with  each  pound  of  dry  air,  pounds, 
ft'fg  =  latent  heat  of  vaporization  at  tl,  Btu  per  pound. 

Since  this  definition  holds  for  any  mixture  of  dry  air  and  water  vapor, 
the  total  heat  of  a  mixture  with  a  relative  humidity  of  100  per  cent  and  at 
a  temperature  equal  to  the  wet-bulb  temperature  (/!)  is 


-  0) 


(11) 


By  equating  Equation  10  to  Equation  11,  the  equation  for  the  adiabatic 
saturation  process,  Equation  9a,  follows.  This  demonstrates  that  the 
adiabatic  saturation  process  at  constant  wet-bulb  temperature  is  also  a 
process  of  constant  total  heat.  In  short,  the  total  heat  of  a  mixture  of  dry 
air  and  water  vapor  is  the  same  for  any  two  states  of  the  mixture  at  the 
same  wet-bulb  temperature.  This  fact  furnishes  a  convenient  means  of 
finding  the  total  heat  of  an  air-vapor  mixture  in  any  state. 

Example  5.  Find  the  total  heat  of  an  air-vapor  mixture  having  a  dry-bulb  tempera- 
ture of  85  F  and  a  wet-bulb  temperature  of  70  F. 

Solution.  From  Table  5,  for  saturation  at  the  wet-bulb  temperature  Wv  =  0.01578, 
and  from  Equation  11, 

Sr  =  Cpa  (70  -  0)  +  0.01578  Wtg  =  16.9  +  16.61  =  33.51 
17 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

By  'considering  the  temperatures  in  Table  5  to  be  wet-bulb  readings,  the 
total  heat  of  any  air-  vapor  mixture  may  be  obtained  from  the  last  column 
in  the  table. 

Enthalpy 

This  total  heat  of  an  air-vapor  mixture  is  not  exactly  equal  to  the  true 
heat  content  or  enthalpy  of  the  mixture  since  the  heat  content  of  the 
liquid  is  not  included  in  Equation  10.  •  With  the  meaning  of  heat  content 
in  agreement  with  present  practise  in  other  branches  of  thermodynamics, 
the  true  heat  content  of  a  mixture  of  dry  air  and  water  vapor  (with  0  F 
as  the  datum  for  dry  air,  and  the  saturated  liquid  at  32  F  as  the  datum 
for  the  water  vapor)  is 

h  =  cPa  (t  -  0)  4-  W  hs  =  0.24  (*  -  0)  +  W  hs  (12) 

where 

h  =  the  heat  content  of  the  mixture,  Btu  per  pound  of  dry  air. 
t  =  the  dry-bulb  temperature,  degrees  Fahrenheit. 
W  =  the  weight  of  vapor  per  pound  of  dry  air,  pounds. 
7fs  =  the  heat  content  of  the  vapor  in  the  mixture,  Btu  per  pound. 

The  heat  content  of  the  water  vapor  in  the  mixture  may  be  found  in 
steam  charts  or  tables  when  the  dry-bulb  temperature  and  the  partial 
pressure  of  the  vapor  are  known.  Or,  since  the  heat  content  of  steam  at 
low  partial  pressures,  whether  super-heated  or  saturated,  depends  only 
upon  temperature,  the  following  empirical  equation,  derived  from 
Keenan's  Steam  Tables,  may  be  used: 

hs  =  1059.2  +  0.45  t  (13) 

Substituting  this  value  of  hs  in  Equation  12,  the  heat  content  of  the 
mixture  is 

h  =  0.24  (t  -  0)  +  W  (1059.2  +  0.45  t}  (14) 

An  energy  equation  can  be  written  that  applies,  in  general,  to  various 
air-conditioning  processes,  and  this  equation  can  be  used  to  determine  the 
quantity  of  heat  transferred  during  such  processes.  In  the  most  general 
form,  this  equation  may  be  explained  with  the  aid  of  Fig.  1  as  follows: 

The  rectangle  may  represent  any  apparatus,  e.g.,  a  drier,  humidifier,  dehumidifier, 
cooling  tower,  or  the  like,  by  proper  choice  of  the  direction  of  the  arrows. 

In  general,  a  mixture  of  air  and  water  vapor,  such  as  atmospheric  air,  enters  the 
apparatus  at  1  and  leaves  at  3.  Water  is  supplied  at  some  temperature,  fe.  For  the  flow 
of  1  Ib  of  dry  air  (with  accompanying  vapor)  through  the  apparatus,  provided  there  is  no 
appreciable  change  in  the  elevation  or  velocity  of  the  fluids  and  no  mechanical  energy 
delivered  to  or  by  the  apparatus, 


or 

Eh  -  Re  =  ^  -  A!  -  (W*  -  Wi)  Jh  (15) 

where 

Eh  -  the  quantity  of  heat  supplied  per  pound  of  dry  air,  Btu. 
j£c  =  the  quantity  of  heat  lost  externally  by  heat  transfer  from  the  ^apparatus, 
Btu  per  pound  of  dry  air. 

18 


CHAPTER  1 — FUNDAMENTALS  OF  HEATING  AND  AIR  CONDITIONING 

Wi  =  the  weight  of  water  vapor  entering,  per  pound  of  dry  air. 
Ws  =  the  weight  of  water  vapor  leaving,  per  pound  of  dry  air. 
fh  —  the  heat  content  of  the  water  supplied  at  t»,  Btu  per  pound. 
hz  —  hi  —  the  increase  'in  the  heat  content  of  the  air- water  vapor  mixture  in  passing 
through  the  apparatus,  Btu  per  pound  of  dry  air 
-  0.24  (fe  -  fr)  -f  Wz  (1059.2  +  0.45  fc)  -  Wl  (1059.2  -f  0.45*0 

The  net  quantity  of  heat  added  to  or  removed  from  air-water  vapor 
mixtures  in  air  conditioning  work  is  frequently  approximated  by  taking 
the  differences  in  total  heat  at  exit  and  entrance. 

For  example,  in  Fig.  1,  an  approximate  result  is 

Eh  -  Re  =  S3  -  Si  (16) 

where 

23  =  the  total  heat  of  the  air-vapor  mixture  at  exit,  Btu  per  pound  of  dry  air. 

Si  —  the  total  heat  of  the  air- vapor  mixture  at  entrance,  Btu  per  pound  of  dry  air. 

From  the  definitions  of  total  heat  and  heat  content,  it  may  be  demon- 
strated that  Equation  16  is  exactly  equivalent  to  Equation  15,  when,  and 
only  when,  ^3  =  t\  —  fe;  i.e.,  when  the  initial  and  final  wet-bulb  tempera- 
tures and  the  temperature  of  the  water  supplied  are  equal.  The  one  pro- 
cess that  meets  these  conditions  is  adiabatic  saturation,  and  either 
equation  will  give  a  result  of  zero;  for  other  conditions,  Equation  16  is 
approximate  'but  satisfactory  for  many  calculations. 
.  The  following  problems  illustrate  the  application  of  these  principles: 

Example  6.  Heating  (data  from  Example  2).  Assuming  the  water  to  be  supplied  at 
50  F,  the  net  quantity  of  heat  supplied  is,  from  Equation  15, 

JSJk  -  jRe  =  0.24  (70  -  0)  +  0.000547  X  0.45  (70  -  0)  -f  0.005633 

or 

1059.2  -f  0.45  X  70  -  (50  -  32)  =  22.87  Btu  per  pound  of  dry  air. 

Example  7.  Cooling  (data  from  Example  3).  If  the  condensate  is  removed  at  54  F 
the  quantity  of  heat  removed  is  found  from  Equation  15,  by  proper  regard  to  the  arrow 
direction  in  Fig.  1, 

Eh  +  J?c  =  0.24  (84  -  54)  -f  0.00887  X  0.45  (84  -  54)  +  0.00358 

or 

1059.2  +  0.45  X  84  -  (54  -  32)  =  11. 17  Btu  per  pound  of  dry  air. 

Using  Table  5,  the  initial  total  heat  of  the  air-vapor  mixture,  since  the  wet-bulb 
temperature  is  70  F,  is  33.51  Btu  per  pound  of  dry  air. 

The  final  total  heat  is,  from  Table  5,  since  the  exit  air  is  saturated,  22.45  Btu  per 
pound.  Hence,  using  Equation  16,  the  quantity  of  heat  removed  is,  approximately, 
(33.51  —  22.45)  or  11.06  Btu  per  pound  of  dry  air.  The  degree  of  approximation  to  the 
correct  result  is  evident  in  this  example. 

PSYCHROMETRIC  CHART4 

The  Bulkeley  Psychrometric  Chart5,  as  revised  will  be  found  as  an 
insert  between  pages  18  and  19.  It  shows  graphically  the  relationships 
expressed  in  Equations  9a  and  9b.  It  also  gives  the  grains  of  moisture  per 


*See  A  Review  of  Psychrometric  Charts,  C.  O.  Mackey  (Heating  and  Ventilating*  June,  July,  1931V 
•The-  Bulkeley  Psychrometric  Chart  was  presented  to  the  Society  in  1926.  ,  (See  A.S.H.V.E.  Tsu 
ACTIONS,  Vol.  32,  1926.)    Single  copy  of  the  chart  can  be  furnished  at  a  cost  of  $  .50. 

19 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


r  r 


Wj  Ib.  Water  Vapor 
1  Ib.  Dry  Air 


V^  Ib.  Water  Vapor 
1  )b.  Dry  Air 


2     \     2 
(W3~Wi)to.  Water 

FIG.  1.    DIAGRAM  ILLUSTRATING  ENERGY  EQUATION  15 


pound  of  dry  air  for  saturation,  the  grains  of  moisture  per  cubic  foot  of 
saturated  air,  the  total  heat  in  Btu  per  pound  of  dry  air  saturated  with 
moisture,  and  the  weight  of  the  dry  air  in  pounds  per  cubic  foot.  ^  Fig.^2 
shows  the  procedure  to  follow  in  using  the  Bulkeley  Chart.  The  directrix 
curves  above  the  saturation  line  are  as  follows: 

A  is  the  total  heat  in  Btu  contained  in  the  mixture  above  0  F,  and  is  to  be  referred 
to  the  column  of  figures  at  the  left  side  of  the  chart.  Heat  of  the  liquid  is  not  included. 

B  is  the  grains  of  moisture  of  water  vapor  contained  in  each  pound  of  the  saturated 
mixture  and  is  to  be  referred  to  the  figures  at  the  left  side  of  the  chart. 

C  is  the  grains  of  moisture  of  water  vapor  per  cubic  foot  of  saturated  mixture,  and  is 
to  be  referred  to  the  figures  at  the  left  side  of  the  chart  which  are  to  be  divided  by  10. 

D  Is  the  weight  in  decimal  fractions  of  a  pound,  of  one  cubic  foot  of  the  saturated 
mixture,  and  is  referred  to  the  first  column  of  figures  to  the  right  of  the  saturation  line 
between  the  vertical  dry-bulb  temperature  lines  170  and  180  F.  The  relative  density  of 


AB-C-D-E*  Directrix  Lines 
D,aL'Dry  Bulb  line 
D.  P.  Lc  Dew  Point  Line 


6.P.LB.=Grains  Moisture  perLb.Drv  AirSahiraM 
T.H.»Totel  Heat  per  Lb.Dry  Air  Saturated 
V.P.= Vapor  Pressure  in  Mm.  Mercury 
6.RCF.S =6roins Moisture  per Cu.Ft  Saturated  Air 


R.H.L=Relative  Humidity  Line 
W.Bl,»WetBulbUne 

S.L"  Saturation Ung 

WJ>C£i=l%TtperCu.Ft,in  Lbs.Saturated 
R-D.S.'Relative  Density  perCu.F-r.Saturcrted 
WP.CF.O.«Retofive  Dererty  per  Cu.Fr.Dry 
R.D.D.=RetfltiveDensfryperCu.Ft.Ory 


6  P  r  F  x  Abs.Temp.gt  P.P. . 
U       Abs.Temp.crtD.B." 
-Abs.Temp.crt  a  P. . 

Abs.Temp,atD.B. 
R  n    y  Abs.Temp.a+P.P.s 
AbsJemp.atD.B. 


*G.P.C.F  at  Partial  Saturation 
W.P.C.F  at  Partial  Saturation 
R.D.  at  Partial  Saturation 


FIG.  2.    DIAGRAMS  SHOWING  PROCEDURE  TO  FOLLOW  IN  USING  BULKELEY  CHART 


20 


CHAPTER  1 — FUNDAMENTALS  OF  HEATING  AND  AIR  CONDITIONING 

the  mixture  is  read  in  a  similar  manner  from  the  same  curve  by  the  column  of  figures 
between  the  vertical  dry-bulb  temperature  lines  180  and  190  F. 

E  is  similar  to  D  but  is  for  dry  air,  devoid  of  all  moisture  or  water  vapor.  For  con- 
venience, the  approximate  absolute  temperature  of  500  F  is  given  at  40  F  on  the  satura- 
tion line  for  the  purpose  of  calculating  volume,  weight  per  cubic  foot,  and  relative  density 
at  partial  saturation. 

METHOD  OF  USING  THE  CHART 

Example  8.  Relative  Humidity:  At  the  intersection  of  the  78  F  wet-bulb  line  and  the 
95  F  dry-bulb  line,  the  relative  humidity  is  read  directly  on  the  straight  diagonal  lines 
as  46  per  cent. 

Example  9.  Dew  Point:  At  the  intersection  of  the  78  F  wet-bulb  line,  the  dew-point 
temperature  is  read  directly  on  the  horizontal  temperature  lines  as  70.9  F. 

Example  10.  Vapor  Pressure:  At  the  intersection  of  the  78  F  wet-bulb  line  and  the 
95  F  dry-bulb  line,  pass  in  a  horizontal  direction  to  the  left  of  the  chart  and  on  the 
logarithmic  scale  read  the  vapor  pressure  as  19.4  millimeters  of  mercury.  (Divide  by 
25.4  for  inches.) 

Example  11.  Total  Heat  Above  0  F  in  Mixture  per  Pou?id  of  Dry  Air  Saturated  with 
Moisture:  From  where  the  wet-bulb  line  joins  the  saturation  line,  pass  in  a  vertical 
direction  on  the  78  F  dry-bulb  line  to  its  intersection  with  curve  A  and  on  the  logarithmic 
scale  at  the  left  of  the  chart  read  40.6*  Btu  per  pound  of  mixture.  The  use  of  this  curve 
to  obtain  the  total  heat  in  the  mixture  at  any  wet-bulb  temperature  is  a  great  con- 
venience, as  the  number  of  Btu  required  to  heat  the  mixture  and  humidify  it,  as  well  as 
the  refrigeration  required  to  cool  and  dehumidify  the  mixture,  can  be  obtained  by 
taking  the  difference  in  total  heat  before  and  after  treatment  of  the  mixture. 

Example  12.  Grains  of  Moisture  per  Pound  of  Mixture:  From  70.9  F  dew-point 
temperature  on  the  saturation  line,  pass  vertically  to  the  intersection  with  curve  B  and 
on  the  logarithmic  scale  at  the  left  read  114  grains  of  moisture  per  pound. 

Example  18.  Grains  of  Moisture  per  Cubic  Foot  of  Mixture,  Partially  Saturated:  From 
70.9  F  dew-point  temperature  on  the  saturation  line  proceed  in  a  vertical  direction  to 
curve  C,  and  on  the  logarithmic  scale  to  the  left  read  83.3  which,  divided  by  10,  gives 
8.33  grains.  A  temperature  of  70.9  F  is  equal  to  an  absolute  temperature  of  530.9,  and 

530  9 
95  F  equals  555,  absolute  temperature.     Therefore,    K  '     X  8.33   =  7.97  grains  per 

ooo 
cubic  foot  of  partially  saturated  mixture. 

Example  14-  Grains  of  Moisture  per  Cubic  Foot  of  Dry  Air,  Saturated:  Starting  at  the 
saturation  line  at  the  desired  temperature,  pass  in  a  vertical  direction  to  curve  C  and  on 
the  logarithmic  scale  at  the  left,  read  a  number  which,  divided  by  10,  will  give  the 
answer. 

Example  15.  Weight  per  Cubic  Foot  of  Dry  Air  and  Relative  Density:  From  the  point 
where,  for  example,  die  70  F  vertical  dry-bulb  line  intersects  curve  E,  pass  to  right  side 
and  read  0.075  Ib ;  if  cubic  feet  per  pound  are  desired,  divide  1  by  this  amount.  The 
relative  density  is  read  immediately  to  the  right  as  1.00. 

Example  16.  Weight  per  Cubic  Foot  of  Saturated  Air  and  Relative  Density:  From  the 
point  where,  for  example,  the  70  F  vertical  line  intersects  the  curve  D,  pass  to  the  right 
and  read  weight  per  cubic  foot  as  0.07316  with  a  relative  density  of  0.9755  for  saturated 
air  at  70  F. 

Example  17.  Weight  per  Cubic  Foot  and  Relative  Density  of  Partially  Saturated  Air: 
Air  at  50  F  and  a  wet-bulb  temperature  of  46  F  is  to  be  heated  to  130  F.  The  wet-  and 
dry-bulb  lines  intersect  at  a  dew-point  temperature  of  42  F.  Pass  to  the  left  where  this 
dew-point  line  intersects  the  saturation  line  and  then  pass  in  a  vertical  direction  to  where 
the  42  F  dry-bulb  line  intersects  with  curve  D.  Then  pass  directly  to  the  right  and  read 
the  weight  per  cubic  foot  of  saturated  air  at  42  F  as  0.07844  and  the  relative  density  as 
1.046.  The  absolute  temperature  at  42  F  is  502,  and  at  130  F  is  590.  Therefore, 

CAO 

*j~   «  0.851.    The  weight  of  1  cu  ft  of  air  at  50  F  dry-bulb  and  46  F  wet-bulb  when 

heated  to  130  F  is  0.07844  X  0.851  =  0.06675,  and  the  relative  density  is  1.046  X  0.851 
=  0.89. 

21 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


PROPERTIES  OF  STEAM 

Steam  is  water  vapor  which  exists  in  the  vaporous  condition  because 
sufficient  heat  has  been  added  to  the  water  to  supply  the  latent  heat  of 
evaporation  and  change  the  liquid  into  vapor.  This  change  in  state  takes 
place  at  a  definite  and  constant  temperature  which  is  determined  solely 
by  the  pressure  of  the  steam.  The  volume  of  a  pound  of  steam  is  the 
specific  wlume  which  decreases  as  the  pressure  increases.  The  reciprocal 
of  this,  or  the  weight  of  steam  per  cubic  foot,  is  the  density.  (See  Table  7.) 

Steam  which  is  in  contact  with  the  water  from  which  it  was  generated  is 
known  as  saturated  steam.  If  it  contains  no  actual  water  in  the  form  of 
mist  or  priming,  it  is  called  dry  saturated  steam.  If  this  be  heated  and  the 
pressure  maintained  the  same  as  when  it  was  vaporized,  its  temperature 
will  increase  and  it  will  become  superheated,  that  is,  its  temperature  will 
be  higher  than  that  of  saturated  steam  at  the  same  pressure. 

PROPERTIES  OF  WATER 

Composition  of  Water.  Water  is  a  chemical  compound  (H20)  formed  by 
the  union  of  two  volumes  of  hydrogen  and  one  volume  of  oxygen,  or  two 
parts  by  weight  of  hydrogen  and  16  parts  by  weight  of  oxygen. 

Density  of  Water.  Water  has  its  greatest  density  at  39.2  F,  and  it 
expands  when  heated  or  cooled  from  this  temperature.  At  62  F  a  U.  S. 
gallon  of  231  cu  in.  of  water  weighs  approximately  8J^  Ib,  and  a  cubic  foot 
of  water  is  equal  to  7.48  gal.  The  specific  volume  of  water  depends  on  the 
temperature  and  it  is  always  the  reciprocal  of  its  density.  (See  Table  8.) 

Water  Pressures.  Pressures  are  often  stated  in  feet  or  inches  of  water 
column.  At  62  F,  with  h  equal  to  the  head  in  feet,  the  pressure  of  a 
column  of  water  is  62.3S3&  Ib  per  square  foot,  or  0.433&  Ib  per  square  inch. 
A  column  of  water  2.309  ft  (27.71  in.)  high  exerts  a  pressure  of  one  pound 
per  square  inch  at  62  F. 

Boiling  Point  of  Water.  The  boiling  point  of  water  varies  with  the 
pressure;  it  is  lower  at  higher  altitudes.  A  change  in  pressure  will  always 
be  accompanied  by  a  change  in  the  boiling  point,  and  there  will  be  a  cor- 
responding change  in  the  latent  heat  of  evaporation.  These  values  are 
given  in  Table  7. 

Specific  Heat.  The  specific  heat  of  water,  or  the  amount  of  heat  (Btu) 
required  to  raise  the  temperature  of  one  pound  of  water  one  degree  Fahren- 
heit, varies  with  the  temperature,  but  it  is  commonly  assumed  to  be 
unity  at  all  temperatures.  Steam  tables  are  based  on  exact  values, 
however.  The  specific  heat  of  ice  at  32  F  is  0.492  Btu  per  pound.  The 
amount  of  heat  required  to  raise  one  pound  of  water  at  32  F  through  a 
known  temperature  interval  depends  on  the  average  specific  heat  for  the 
temperature  range. 

Sensible  and  Latent  Heat.  The  heat  necessary  to  raise  the  temperature 
of  one  pound  of  water  from  32  F  to  the  boiling  point  is  known  as  the  heat 
of  the  liquid  or  sensible  heat.  When  more  heat  is  added,  the  water  begins 
to  evaporate  and  expand  at  constant  temperature  until  the  water  is 
entirely  changed  into  steam.  The  heat  thus  added  is  known  as  the  latent 
heat  of  evaporation. 

22 


CHAPTER  1  —  FUNDAMENTALS  OF  HEATING  AND  AIR  CONDITIONING 

TABLE 

Abt.  Pres*.  Temp. 

7.    PROPERTIES  OF  SATURATED  STEAM:    PRESSURE  TABLE** 

Specific  Volume                          Total  Heat                             Entropy 
Sat.                            Sat.                    Sat.                           Sat.              Sat.                             Sat.     Ab«.  Pre**. 

Lb./Sq.  In.  Deg.  F. 

Liquid 

Evap. 

Vapor 

Liquid 

Evap. 

Vapor 

Liquid 

Evap. 

Vapor  Lb./Sq.  In. 

^P 

t 

Vf 

Vfg 

Vg 

hf 

hfg 

kg 

Sf 

Sfg 

Sg 

P 

58.83 

0.01603 

1256.9    1 

,256.9 

26.88 

1058.8 

1085.7 

0.0533 

2.0422 

2.0955 

V&"  &£ 

3  /  n  gw 

70.44 

0.01605 

856.5 

856.5 

38.47 

1052.5 

1091.0 

0.0754 

1.9856 

2.0609 

3/4"  Hg 

r'Hg 

79.06 

0.01607 

652.7 

652.7 

47.06 

1047.8 

1094.9 

0.0914 

1.9451 

2.0365 

l"Hg 

91.75 

0.01610 

4453 

4453 

59.72 

1040.8 

1100.6 

0.1147 

1.8877 

2.0024 

iVi"  s& 

2"Hg 

101.17 

0.01613 

339.5 

339.5 

69.10 

1035.7 

1104.S 

0.1316 

1.846S 

1.9784 

2"Hg 

2W'  Hg 

108.73 

0.01616 

275.2 

2752 

76.63 

1031.5 

1108.1 

0.1450 

1.8148 

1.9598 

2V2"Hg 

3"Hg 

115.08 

0.01618 

231.8 

231.8 

82.96 

1027.9 

1110.8 

0.1551 

1.7885 

1.9446 

3"Hg 

1.0 

101.76 

0.01614 

333.8 

333.9 

69.69 

10353 

1105.0 

0.1326 

1.8442 

1.9769 

1.0 

2.0 

126.10 

0.01623 

173.94 

173.96 

93.97 

1021.6 

1115.6 

0.1750 

1.7442 

1.9192 

2.0 

3.0 

141.49 

0.01630 

118,84 

118.86 

10933 

1012.7 

1122.0 

0.2009 

1.6847 

1.8856 

3.0 

4.0 

152.99 

0.01636 

90.72 

90.74 

120.83 

1005.9 

1126.8 

0.2198 

1.6420 

1.8618 

4.0 

5.0 

162.25 

0.01641 

73.59 

73.61 

130.10 

1000.4 

1130.6 

0.2348 

1.6088 

1.8435 

5.0 

6.0 

170.07 

0.01645 

62.03 

62.05 

137.92 

995.8 

1133.7 

0.2473 

1.5814 

1.8287 

6.0 

7.0 

176.85 

0.01649 

53.68 

53.70 

144.71 

991.7 

1136.4 

0.2580 

1.5582 

1.8162 

7.0 

8.0 

182.87 

0.01652 

4738 

4739 

150.75 

988.1 

1138.9 

0.2674 

1.5379 

1.8053 

8.0 

9.0 

188.28 

0.01656 

42.42 

42.44 

156.19 

984.8 

1141.0 

0.2758 

1.5200 

1.7958 

9.0 

10.0 

193.21 

0.01658 

38.44 

38.45 

161.13 

981.8 

1143.0 

0.2834 

1.5040 

1.7874 

10.0 

11.0 

197.75 

0.01661 

35.15 

35.17 

165.68 

979.1 

1144.8 

0.2903 

1.4894 

1.7797 

11.0 

12.0 

201.96 

0.01664 

32.40 

3Z.42 

169.91 

976.5 

1146.4 

0.2968 

1.4760 

1.7727 

12.0 

13.0 

205.88 

0.01666 

30.06 

30.08 

173.85 

974.1 

1147.9 

0.3027 

1.4636 

1.7663 

13.0 

14.0 

209.56 

0.01669 

28.05 

28.06 

177.55 

971.8 

11493 

03082 

1.4521 

1.7604 

14.0 

14.696 

212.00 

0.01670 

26,80 

26.82 

180.00 

970.2 

1150.2 

03119 

1,4446 

1.7564 

14.696 

16.0 

21632 

0.01673 

24.75 

24.76 

18435 

967.4 

1151.8 

03184 

1.4312 

1.7496 

16.0 

18.0 

222.40 

0.01678 

22.16 

22.18 

190.48 

963.5 

1154.0 

03274 

1.4127 

1.7402 

18.0 

20.0 

227.96 

0.01682 

20.078 

20.095 

196.09 

959.9 

1155.0 

03356 

13960 

1.7317 

20.0 

22.0 

233.07 

0.01685 

18363 

18380 

201.25 

956.6 

1157.8 

03431 

13809 

1.7240 

22.0 

24.0 

237.82 

0.01689 

16.924 

16.941 

206.05 

953.4 

1159.5 

03500 

13670 

1.7170 

24.0 

26.0 

242.25 

0.01692 

15.701 

15.718 

210.54 

950.4 

1161.0 

J03564 

13542 

1.7106 

26.0 

28.0 

246.41 

0.01695 

14.647 

14.664 

214.75 

947.7 

1162.4 

03624 

13422 

1.7046 

28.0 

30.0 

25034 

0.01698 

13.728 

13.745 

218.73 

945.0 

1163.7 

03680 

13310 

1.6990 

30.0 

32.0 

254.05 

0.01701 

12.923 

12.940 

222.50 

942.5 

1165.0 

03732 

13206 

1.6938 

32.0 

34.0 

257.58 

0.01704 

12.209 

12.226 

226.09 

940.0 

1166.1 

03783 

13107 

1.6890 

34.0 

36.0 

260.94 

0.01707 

11.570 

11.587 

229.51 

937.7 

1167.2 

03830 

13014 

1.6844 

36.0 

38.0 

264.16 

0.01710 

10.998 

11.015 

232.79 

935.5 

11683 

03876 

1.2925 

1.6800 

38.0 

40.0 

267.24 

0.01712 

10.480 

10.497 

235.93 

9333 

1169.2 

03919 

1.2840 

1.6759 

40.0 

42.0 

270.21 

0.01715 

10.010 

10.027 

238.95 

931.2 

1170.2 

03961 

1.2759 

1.6720 

42.0 

44.0 

273.06 

0.01717 

•       9.582 

9.599 

241.86 

929.2 

1171.1 

0.4000 

1.2682 

1.6683 

44.0 

46.0 

275.81 

0.01719 

9.189 

9.207 

244.67 

9272 

1171.9 

0.4039 

1.2608 

1.6647 

46.0 

48.0 

278.45 

0.01722 

8.829 

8.846 

24737 

925.4 

1172.7 

0.4076 

1.2537 

1.6613 

48.0 

50.0 

281.01 

0.01724 

8.496 

8.514 

249.98 

923.5 

1173.5 

0.4111 

1.2469 

1.6580 

60.0 

52.0 

283.49 

0.01726 

8.189 

8.206 

252.52 

921.7 

11743 

0.4145 

1.2404 

1.6549 

62.0 

54.0 

285,90 

0.01728 

7.902 

7.919 

254.99 

920.0 

1175.0 

0.4178 

1.2340 

1.6518 

64.0 

56.0 

288.23 

0.01730 

7.636 

7.653 

25738 

9183 

1175.7 

0.4210 

1.2279 

1.6489 

66.0 

58.0 

290.50 

0.01732 

7388 

7.405 

259.71 

916.6 

1176.4 

0.4241 

1.2220 

1,6461 

68.0 

60.0 

292.71 

0.01735 

7.155 

7.172 

261.98 

915,0 

1177.0 

0.4271 

1.2162 

1.6434 

60.0 

62.0 

294.85- 

0.01737 

6.937 

6.955 

264.18 

913.4 

1177.6 

0.4300 

1.2107 

1.6407 

62.0 

64.0 

296.94 

0.01739 

6.732 

6.749 

26633 

911.9 

1178.2 

0.4329 

1.2053 

1.6382 

64.0 

66.0 

298.98 

0.01741 

6.539 

6.556 

268.43 

910.4 

1178.8 

0.4356 

1.2001 

1.6357 

66.0 

68.0 

300.98 

0.01743 

6357 

6375 

270.49 

908.9 

1179,4 

0.4384 

1.1950 

1.6333 

68.0 

70.0 

302.92 

0.01744 

6.186 

6.203 

272.49 

907.4 

1179.9 

0.4410 

1.1900 

1.6310 

70.0 

72.0 

304.82 

0.01746 

6.024 

6.041 

274.45 

906.0 

1180.5 

0.4435 

1.1852 

1.6287 

72.0 

74.0 

306.68 

0.01748 

5.870 

5.887 

27637 

904.6 

1181.0 

0.4460 

1.1805 

1.6265 

74.0 

76.0 

30830 

0.01750 

5.723 

5.741 

278.25 

903.2 

1181.5 

0.4485 

1.1759 

1.6244 

76.0 

78.0 

310.28 

.0.01752 

5.584 

5.602 

280.09 

901.9 

1182.0 

0.4509 

1.1714 

1.6223 

78.0 

80.0 

312.03 

0.01754 

5.452 

5.470 

281.90 

900.5 

1182.4 

0.4532 

1.1670 

1.6202 

80.0 

82.0 

313.74 

0.01756 

5325 

5343 

283.67 

899.2 

1182.9 

0.4555 

1.1627 

1.6182 

82.0 

84.0 

315.42 

0.01757 

5.204 

5.222 

285.42 

897.9 

1183.4 

0.4578 

1.1586 

1.6163 

84.0 

86.0 

317.06 

0.01759 

5.089 

5.107 

287.13 

896.7 

1183.8 

0.4599 

1.1545 

1.6144 

86.0 

88.0 

318.68 

0.01761 

4.979 

4.997 

288.80 

895.4 

1184.2 

0-4621 

1.1505 

1.6126 

88.0 

90.0 

320.27 

0.01763 

4.874 

4.892 

290.45 

894.2 

1184.6 

0.4642 

1.1465 

1.6107 

90.0 

92.0 

321.83 

0.01764 

4.773 

4.791 

292.07 

893.0 

1185.0 

0.4663 

1.1427 

1.6090 

92.0 

94.0 

32337 

0.01766 

4.676 

'4.694 

293.67 

891.8 

1185.4 

0.4683 

1.1389 

1,6072 

94.0 

96.0 

324.88 

0.01768 

4.584 

4.602 

•295.25 

890.6 

1185.8 

0.4703 

1.1352 

1.6055 

96.0 

98.0 

32637 

0.01769 

4-494 

4.512 

•JJ96.80 

889.4 

1186.2 

0.4723 

1.1316 

1.6038 

98.0 

•Abstracted  from  Steam  Tables  and  Mottier  Diagram,  by  Prof.  J.  H.  Keenan,  1930  edition,  by  permission 
of  the  publisher,  The  American  Society  of  Mechanical  Engineers. 

23 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


TABLE  7.    PROPERTIES  OF  SATURATED  STEAM:    PRESSURE  TABLE — (Continued) 


Specific  Volume 

Total  Heat 

Entropy 

Abs.  Press. 

Temp. 

Sat. 

Sat. 

Sat. 

Sat. 

Sat. 

Sat. 

Aba.  Pr«ss. 

Lb./Sq.  In. 

DC*.*- 

Liquid 

Evap. 

Vapor 

Liquid 

Evap. 

Vapor 

Liquid 

Evap. 

Vapor 

Lb./Sq.  In. 

P 

t 

Vf 

Vfg 

Vg 

hf 

hfg 

he 

Sf 

Sfg 

Sg 

P 

100.0 

327.83 

0.01771 

4.408 

4.426 

29833 

888.2 

1186.6 

0.4742 

1.1280 

1.6022 

100.0 

102.0 

329.27 

0.01773 

4326 

4344 

299.83 

887.1 

1186.9 

0.4761 

1.1245 

1.6006 

102.0 

104.0 

330.68 

0.01774 

4.247 

4.265 

301.30 

886.0 

11873 

0.4779 

1.1211 

1.5990 

104.0 

106.0 

332.08 

0.01776 

4.171 

4.189 

302.76 

884.9 

1187.6 

0.4798 

1.1177 

1.5974 

106.0 

108.0 

333.44 

0.01777 

4.097 

4.115 

304.19 

883.8 

1188.0 

0.4816 

1.1144 

1.5959 

108.0 

110.0 

334.79 

0.01779 

4.026 

4.044 

305.61 

882.7 

11883 

0.4834 

1.1111 

1.5944 

110.0 

112.0 

336.12 

0.01780 

3.958 

3.976 

307.00 

881.6 

1188.6 

0.4851 

1.1079 

1.5930 

112.0 

114.0 

337.43 

0.01782 

3.892 

3.910 

308.36 

880.6 

1188.9 

0.4868 

1.1048 

1.5915 

114.0 

116.0 

338.72 

0.01783 

3.828 

3.846 

309.71 

879.5 

1189.2 

0.4885 

1.1017 

1.5901 

116.0 

118.0 

340.01 

0.01785 

3.766 

3.784 

311.05 

878.5 

1189.5 

0.4901 

1.0986 

1.5887 

118.0 

120.0 

341.26 

0.01786 

3.707 

3.725 

312.37 

877.4 

1189.8 

0.4918 

1.0956 

1.5874 

120.0 

122.0 

342.50 

0.01788 

3.652 

3.670 

313.67 

876.4 

1190.1 

0.4934 

1.0926 

1.5860 

122.0 

124.0 

343.73 

0.01789 

3.597 

3.615 

314.96 

875.4 

1190.4 

0.4950 

1.0897 

1.5847 

124.0 

126.0 

344.94 

0.01791 

3.542 

3.560 

316.23 

874.4 

1190.6 

0.4965 

1.0868 

1.5834 

126.0 

128.0 

346.14 

0.01792 

3.487 

3.505 

317.49 

873.4 

1190.9 

0.4981 

1.0840 

1.5821 

128.0 

130.0 

347.31 

0.01794 

3.433 

3.451 

318.73 

872.4 

1191.2 

0.4996 

1.0812 

1.5808 

130.0 

132.0 

348.48 

0.01795 

3.383 

3.401 

319.95 

871.5 

1191.4 

0.5011 

1.0784 

1.5796 

132.0 

134.0 

349.64 

0.01796 

3335 

3353 

321.17 

870.5 

1191.7 

0.5026 

1.0757 

1.5783 

134.0 

136.0 

350.78 

0.01798 

3.288 

3306 

32237 

869.6 

1191.9 

0.5041 

1.0730 

1.5771 

136.0 

138.0 

351.91 

0.01799 

3.242 

3.260 

323.56 

868.6 

1192.2 

0.5056 

1.0703 

1.5759 

138.0 

140.0 

353.03 

0.01801 

3.198 

3.216 

324.74 

867.7 

1192.4 

0.5070 

1.0677 

1.5747 

140.0 

142.0 

354.14 

0.01802 

3.155 

3.173 

325.91 

.866.7 

1192.6 

0.5084 

1.0651 

1.5735 

142.0 

144.0 

355.22 

0.01804 

3.112 

3.130 

327.06 

865.8 

1192.9 

0.5098 

1.0625 

1.5724 

144.0 

146.0 

35631 

0.01805 

3.071 

3.089 

328.20 

864-9 

1193.1 

0.5112 

1.0600 

1.5712 

146.0 

148.0 

357.37 

0.01806 

3.031 

3.049 

32932 

864.0 

11933 

0.5126 

1.0575 

1.5701 

148.0 

150.0 

358.43 

0.01808 

2.992 

3.010 

330.44 

863.1 

1193.5 

0.5140 

1.0550 

1.5690 

150.0 

152.0 

359.47 

0.01809 

2.954 

2.972 

331.54 

862.2 

1193.7 

0.5153 

1.0526 

1.5679 

152.0 

154.0 

360.51 

0.01810 

2.917 

2.935 

332.64 

8613 

1193.9 

0.5166 

1.0502 

1.5668 

154.0 

156.0 

361.53 

0.01812 

2.882 

2.9QO 

333.72 

860.4 

1194.1 

0.5180 

1.0478 

1.5658 

156.0 

158.0 

362.54 

0.01813 

2.846 

2.864 

334.80 

859.5 

11943 

0.5193 

1.0454 

1.5647 

158.0 

160.0 

363.55 

0.01814 

2.812 

2.830 

335.86 

858.7 

1194.5 

0.5205 

1.0431 

1.5636 

160.0 

162.0 

364.54 

0.01816 

2.779 

2.797 

336.91 

857.8 

1194.7 

0.5218 

1.0408 

1.5626 

162.0 

164.0 

365.52 

0.01817 

2.746 

2.764 

337.95 

857.0 

1194.9 

0.5230 

1.0385 

1.5616 

164.0 

166.0 

366.50 

0.01818 

2.715 

2.733 

338.99 

856.1 

1195.1 

0.5243 

1.0363 

1.5606 

166,0 

168.0 

367.46 

0.01819 

2.683 

2.701 

340.01 

855.2 

11953 

0.5255 

1.0340 

1.5596 

168.0 

170.0 

368.42 

0.01821 

2.653 

2.671 

341.03 

854.4 

1195.4 

0.5268 

1.0318 

1.5586 

170.0 

172.0 

369.37 

0.01822 

2.623 

2.641 

342.04 

853.6 

1195.6 

0.5280 

1.0296 

1.5576 

172.0 

174.0 

37031 

0.01823 

2.594 

2.612 

343.04 

852.7 

1195.8 

0.5292 

1.0275 

1.5566 

174.0 

176.0 

371.24 

0.01825 

2.566 

2.584 

344.03 

851.9 

1196.0 

0.5304 

1.0253 

1.5557 

176.0 

178.0 

372.16 

0.01826 

2.538 

2.556 

345.01 

851.1 

1196.1 

0.5315 

1.0232 

1.5548 

178.0 

180.0 

373.08 

0.01827 

2.511 

2.529 

345.99 

850.3 

1196.3 

0.5327 

1.0211 

1.5538 

180.0 

182.0 

374.00 

0.01828 

2.484 

2.502 

346.97 

849.5 

1196.4 

0.5339 

1.0190 

1.5529 

182.0 

184.0 

374.90 

0.01829 

2.458 

2.476 

347.94 

848.6 

1196.6 

0.5350- 

1.0169 

1.5520 

184.0 

186.0 

375.78 

0.01831 

2.433 

2.451 

348.89 

847.9 

1196.8 

0.5362 

1.0149 

1.5511 

186,0 

188.0 

376.67 

0.01832 

2.407 

2.425 

349.83 

847.1 

1196.9 

0.5373 

1.0129 

1.5502 

188.0 

190.0 

377.55 

0.01833 

2383 

2.401 

350.77 

846.3 

1197.0 

0.5384 

1.0109 

1.5493 

190.0 

192.0 

378.42 

0.01834 

2359 

2377 

351.70 

845.5 

1197.2 

0.5395 

1.0089 

1.5484 

192.0 

194.0 

379.27 

0.01835 

2335 

2353 

352.61 

844.7 

1197.3 

0.5406 

1.0070 

1.5475 

194.0 

196.0 

380.13 

0.01837 

2.312 

2330 

353.53 

844.0 

1197.5 

0.5417 

1.0050 

1.5467 

196.0 

198.0 

380.97 

0.01838 

2.289 

2307 

354.43 

843.2 

1197.6 

0.5427 

1.0031 

1.5458 

198.0 

200.0 

381.82 

0.01839 

2.267 

2.285 

35533 

842.4 

1197.8 

0.5438 

1.0012 

1.5450 

200.0 

205.0 

383.89 

0.01842 

2.213 

2.231 

357.56 

840.5 

1198.1 

0.5465 

0.9964 

1.5429 

205.0 

210.0 

385.93 

0.01844 

2.162 

2.180 

359.76 

838.6 

1198.4 

0.5491 

0.9918 

1.5409 

210.0 

215.0 

387.93 

0.01847 

2.113 

2.131 

361.91 

836.8 

1198.7 

0.5516 

0,9873 

1.5389 

215.0 

220.0 

389.89 

0.01850 

2.066 

2.084 

364.02 

835.0 

1199.0 

0.5540 

0.9829 

1.5369 

220.0 

225.0 

391.81 

0.01853 

2.0208 

2.0393 

366.10 

833.2 

1199.3 

0.5565 

0.9786 

1.5350 

225.0 

230.0 

393.70 

0.01856 

1.9778 

1.9964 

368.14 

831.4 

1199.6 

0.5588 

0.9743 

1.5332 

230.0 

235.0 

395.56 

0.01859 

1.9367 

1.9553 

370.15 

829.7 

1199.8 

0.5612 

0.9702 

1.5313 

235.0 

240.0 

397.40 

0.01861 

1.8970 

1.9156 

372.13 

827.9 

1200.1 

0.5635 

0,9661 

1.5295 

240.0 

245.0 

399.20 

0.01864 

1.8589 

1.8775 

374.09 

826.2 

1200,3 

0,5658 

0.9620 

1.5278 

245.0 

24 


CHAPTER  1  —  FUNDAMENTALS  OF  HEATING  AND  AIR  CONDITIONING 

TABLE  7.    PROPERTIES  OF  SATURATED  STEAM: 

PRESSURE  TABLE  —  (Continued) 

Specific  Volume 

Total  Heat 

Entropy 

Abs.  Press. 
Lb./Sq.  In. 

Temp. 
Dee.  F. 

Sat. 
Liquid 

Evap. 

Sat. 
Vapor 

Sat. 
Liquid 

Evap. 

Sat. 
Vapor 

Sat. 
Liquid 

Evap. 

Sat. 
Vapor 

Abs.  Press. 
Lb./Sq.  In. 

P 

t 

Vf 

VfK 

Vg 

hf 

hfg 

fcg 

Sf 

Sfg 

sg 

P 

250.0 

400.97 

0.01867 

1.8223 

1.8410 

376.02 

824.5 

1200.5 

0.5680 

0.9581 

1.5261 

250.0 

260.0 

404.43 

0.01872 

1.7536 

1.7723 

379.78 

821.2 

1201.0 

0.5723 

0.9504 

1.5227 

260.0 

270.0 

407.79 

0.01877 

1.6895 

1.7083 

383.44 

818.0 

1201.4 

0.5765 

0.9430 

1.5194 

270.0 

280.0 

411.06 

0.01882 

1.6302 

1.6490 

387.02 

814.7 

1201.8 

0.5805 

0.9357 

1.5163 

280.0 

290.0 

414.24 

0.01887 

1.5745 

1.5934 

390.50 

811.6 

1202.1 

0.5845 

0.9287 

1.5132 

290.0 

300.0 

41733 

0.01892 

1.5225 

1.5414 

393.90 

808.5 

1202.4 

0.5883 

0.9220 

1.5102 

300.0 

320.0 

423.29 

0.01901 

1.4279 

1.4469 

400.47 

802.5 

1203.0 

0.5957 

0.9089 

1.5046 

320.0 

340.0 

428.96 

0.01910 

13439 

13630 

406.75 

796.6 

1203.4 

0.6027 

0.8965 

1.4992 

340.0 

360.0 

434.39 

0.01918 

1.2689 

1.2881 

412.80 

790.9 

1203.7 

0.6094 

0.8846 

1.4940 

360.0 

380.0 

439.59 

0.01927 

1.2015 

1.2208 

418.61 

7853 

1203.9 

0.6157 

0.8733 

1.4891 

360.0 

400.0 

444.58 

0.0194 

1.1407 

1.1601 

424.2 

779.8 

1204.1 

0.6218 

0.8625 

1.4843 

400.0 

420.0 

44938 

0.0194 

1.0853 

1.1047 

429.6 

774.5 

1204.1 

0.6277 

0.8520 

1.4798 

420.0 

440.0 

454.01 

0.0195 

1.0345 

1.0540 

434.8 

7693 

1204.1 

0.6334 

0.8420 

1.4753 

440.0 

460.0 

458.48 

0.0196 

0.9881 

1.0077 

439.9 

764.1 

1204.0 

0.6388 

0.8322 

1.4711 

460.0 

480.0 

462.80 

0.0197 

0.9456 

0.9633 

444.9 

759.0 

1203.9 

0.6441 

0.8228 

1.4670 

480.0 

600.0 

466.99 

0.0198 

0.9063 

0.9261 

449.7 

754.0 

1203.7 

0.6493 

0.8137 

1.4630 

500.0 

520.0 

471.05 

0.0198 

0.8701 

0.8899 

454.4 

749.0 

1203.5 

0.6543 

0.8048 

1.4591 

520.0 

640.0 

474.99 

0.0199 

0.8363 

0.8562 

459.0 

744.1 

1203.2 

0.6592 

0.7962 

1.4554 

540.0 

560.0 

478.82 

0.0200 

0.8047 

0.8247 

463.6 

7393 

1202.9 

0.6639 

0.7878 

1.4517 

560.0 

680.0 

482.55 

0.0201 

0.7751 

0.7952 

468.0 

734.5 

1202.5 

0.6686 

0.7796 

1.4482 

580.0 

600.0 

486.17 

0.0202 

0.7475 

0.7677 

4723 

729.8 

1202.1 

0.6731 

0.7716 

1.4447 

600.0 

620.0 

489.71 

0.0202 

0.7217 

0.7419 

476.6 

725.1 

1201.7 

0.6775 

0.7638 

1.4413 

620.0 

640.0 

493.16 

0.0203 

0.6972 

0.7175 

480.8 

720.5 

1201.2 

0.6818 

0.7562 

1.4380 

640.0 

660.0 

496.53 

0.0204 

0.6744 

0.6948 

484.9 

715.9 

1200.8 

0.6861 

0.7487 

1.4348 

660.0 

680.0 

499.82 

0.0205 

0.6527 

0.6732 

488.9 

7113 

1200.2 

0.6902 

0.7414 

1.4316 

680.0 

700.0 

503.04 

0.6206 

0.6321 

0.6527 

492.9 

706.8 

1199.7 

0.6943 

0.7342 

1.4285 

700.0 

720.0 

506.19 

0.0206 

0.6128 

0.6334 

496.8 

702.4 

1199.2 

0.6983 

0.7272 

1.4255 

720.0 

740.0 

509.28 

0.0207 

0.5944 

0.6151 

500.6 

697.9 

1198.6 

0.7022 

0.7203 

1.4225 

740.0 

760.0 

51230 

0.0208 

0.5769 

0.5977 

504.4 

693.5 

1198.0 

0.7060 

0.7136 

1.4196 

760.0 

780.0 

515.27 

0.0209 

0.5602 

0.5811 

508.2 

689.2 

1197.4 

0.7098 

0.7069 

1.4167 

780.0 

SOO.O 

518.18 

0.0209 

0.5444 

0.5653 

511.8 

684.9 

1196.7 

0.7135 

0.7004 

1.4139 

800.0 

820.0 

521.03 

0.0210 

0.5293 

0.5503 

515.5 

680.6 

1196.0 

0.7171 

0.6940 

1.4111 

820.0 

840.0 

523.83 

0.0211 

0.5149 

0.5360 

519.0 

676.4 

1195.4 

0.7207 

0.6877 

1.4084 

840.0 

860.0 

526.58 

0.0212 

0.5013 

0.5225 

522.6 

672.1 

1194.7 

0.7242 

0.6815 

1.4057 

860.0 

880.0 

529.29 

0.0213 

0.4881 

0.5094 

526.0 

667.9 

1194.0 

0.7277 

0.6754 

1.4031 

880.0 

900.0 

531.95 

0.0213 

0.4756 

0.4969 

529.5 

663.8 

11933 

0.7311 

0,6694 

1.4005 

900.0 

920.0 

534.56 

0.0214 

0.4635 

0.4849 

532.9 

659.7 

1192.6 

0.7344 

0.6635 

13980 

920.0 

940.0 

537.13 

0.0215 

0.4520 

0.4735 

536.2 

655.6 

1191.8 

0.7377 

0.6577 

13954 

940.0 

960.0 

539.66 

0.0216 

0.4409 

0.4625 

539.6 

651.5 

1191.1 

0.7410 

0.6520 

13930 

960.0 

980.0 

542.14 

0.0217 

0.4303 

0.4520 

542.8 

647.5 

11903 

0.7442 

0.6464 

13905 

980.0 

1000.0 

544.58 

0.0217 

0.4202 

0.4419 

546.0 

643.5 

1189.6 

0.7473 

0.6408 

13881 

1000.0 

1050.0 

550.53 

0.0219 

03960 

0.4179 

554.0 

633.6 

1187.6 

0.7550 

0.6273 

13822 

1050.0 

1100.0 

556.28 

0.0222 

03738 

03960 

561.7 

623.9 

1185.6 

0.7624 

0.6141 

13765 

1100.0 

1150.0 

561.81 

0.0224 

03540 

03764 

569.2 

6143 

11835 

0.7695 

0.6014 

13709 

1150.0 

1200.0 

567.14 

0.0226 

03356 

03582 

5763 

604.9 

1181.4 

0.7764 

0.5891 

13656 

1200.0 

1250.0 

57230 

0.0228 

03187 

03415 

583.6 

595.6 

1179.2 

0.7831 

0.5772 

13603 

1250.0 

1300.0 

57732 

0.0230 

03029 

03259 

590.6 

5863 

1177.0 

0.7897 

0.5654 

13552 

1300.0 

1350.0 

582.21 

0.0232 

0.2884 

03116 

597.5 

577.2 

1174.7 

0.7962 

0.5540 

13501 

1350.0 

1400.0 

586.96 

0.0235 

0.2748 

0.2983 

6043 

568.1 

1172.4 

0.8024 

0.5428 

13452 

1400.0 

1450.0 

591.58 

0.0237 

0.2621 

0.2858 

•611.0 

559.1 

1170.0 

0.8086 

0.5318 

13404 

1450.0 

1500.0 

596.08 

0.0239 

0.2502 

0.2741 

617.5 

550.2 

1167.6 

0.8146 

0.5212 

13357 

1600.0 

1600.0 

604.74 

0.0244 

0.2284 

0.2528 

630.2 

532.6 

1162.7 

0.8262 

0.5003 

13265 

1600.0 

1700.0 

612.98 

0.0249 

0.2089 

0.2338 

642.5 

515.0 

1157.5 

0.8373 

0.4801 

13174 

1700.0 

1800.0 

620,86 

0.0254 

0.1913 

0.2167 

654.7 

497.2 

115L8 

0.8482 

0.4601 

13083 

1800.0 

1900.0 

62839 

0.0260 

0.1754 

0.2014 

666.8 

478,9 

1145.7 

0.8589 

0.4402 

1.2990 

1900.0 

2000.0 

635.6 

0.0265 

0.1610 

0.1875 

679.0 

460.0 

1139.0 

0.8696 

0.4200 

1.2896 

2000.0 

2200.0 

649.2 

0.0277 

0.1346 

0.1623 

703.7 

420.0 

1123.8 

0.8912 

03788 

1.2700 

2200.0 

2400.0 

661.9 

0.0292 

o.im 

0.1404 

729.4 

376.4 

1105.8 

0.9133 

03356 

1.2488 

2400.0 

2600.0 

673.S 

0.0310 

0.0895 

0.1205 

756.7 

327.8 

1084.5 

0.9364 

0.2892 

1.2257 

2600.0 

2800.0 

684.9 

0.0333 

0.0688 

0.1021 

786.7 

2723 

1058.9 

0.9618 

0.2379 

1.1996 

2800.0 

3000.0 

695.2 

0.0367 

0.0477 

0.0844 

823.1 

202.5 

1025.6 

0.9922 

0.1754 

1.1676 

3000.0 

3200.0 

704.9 

O.C459 

0.0142 

0.0601 

887.0 

75.9 

962.9 

1.0461 

0.0651 

1.1112 

3200.0 

3226.0 

706,1 

0.0522 

0 

0.0522 

925.Q 

0 

925.0 

1.0785 

0 

1.0785 

3226.0 

25 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


RATE  OF  EVAPORATION 

In  problems  of  air  conditioning  and  drying,  as  well  as  in  other  industrial 
applications  of  evaporation,  such  as  cooling  towers,  it  is  desirable  to 
determine  the  rate  of  evaporation.  There  are  two  distinct  cases  of 
evaporation.  The  first  case  is  that  in  which  the  source  of  heat  is  primarily 
from  the  water  itself  and  in  which  the  air  temperature  may  even  be  raised. 


TABLE  8.    THERMAL  PROPERTIES  OF  WATER 


TEMPERATURE 
DBG? 

SAT.  PRESS. 
LB  PER  SQ  IN. 

VOLUME  Cu  FT 

PERLB 

WEIGHT  LB  PER 
CuFT 

SPECIFIC 
HEAT 

32 

0.0887 

0.01602 

62.42 

1.0093 

40 

0.1217 

0.01602 

62.42 

1.0048 

50 

0.1780 

0.01602 

62.42 

1.0015 

60 

0.2561 

0.01603 

62.38 

0.9995 

70 

0.3628 

0.01605 

62.31 

0.9982 

80 

0.5067 

0.01607 

62.23 

0.9975 

90 

0.6980 

0.01610 

62.11 

0.9971 

100 

0.9487 

0.01613 

62.00 

0.9970 

110 

1.274 

0.01616 

61.88 

0.9971 

120 

1.692 

0.01620 

61.73 

0.9974 

130 

2.221 

0.01625 

61.54 

0.9978 

140 

2.887 

0.01629 

61.39 

0.9984 

150 

3.716 

0.01634 

61.20 

0.9990 

160 

4.739 

0.01639 

61.01 

0.9998 

170 

5.990 

0.01645 

60.79 

1  .0007 

180 

7.510 

0.01650 

60.61 

1.0017 

190 

9.336 

0.01656 

60.39 

1.0028 

200 

11.525 

0.01663 

60.13 

1.0039 

210 

14.123 

0.01669 

59.92 

1.0052 

212 

14.696 

0.01670 

59.88 

1.0055 

220 

17.188 

0.01676 

59.66 

1.0068 

240 

24.97 

0.01690 

59.17 

1.0104 

260 

35.43 

0.01706 

58.62 

1.0148 

280 

49.20 

0.01723 

58.04 

1.020 

300 

67.01 

0.01742 

57.41 

1.026 

350 

134.62 

0.01797 

55.65 

1.044 

400 

247.25 

0.01865 

53.62 

1.067 

450 

422.61 

0.0195 

51.3 

1.095 

500 

681.09 

0.0205 

48.8 

1.130 

550 

1045.4 

0.0219 

45.7 

1.200 

600 

1544.6 

0.0241 

41.5 

1.362 

700 

3096.4 

0.0394 

25.4 



The  second  is  that  in  which  the  heat  for  evaporation  is  obtained  entirely 
from  the  air  itself,  in  which  case  the  air  is  cooled  and  the  temperature  oJ 
the  water  remains  substantially  constant  at  the  wet-bulb  temperature 
Both  cases,  however,  may  be  reduced  to  a  common  basis  of  calculation 
It  has  been  found  that  the  increase  in  the  rate  of  evaporation  is  nearly  ir 
direct  proportion  to  the  increase  in  the  air  velocity,  and  that  it  is  in  dired 
proportion  to  the  difference  in  vapor  pressure  between  the  vapor  pressure 
of  the  water  and  the  pressure  of  the  vapor  in  the  air. 


26 


CHAPTER  1 — FUNDAMENTALS  OF  HEATING  AND  AIR  CONDITIONING 

The  general  formula  covering  the  experimental  data  may  be  expressed 
as  follows: 

^  =  (a  +  bu)  («'  -  e)  (17) 

where 

dw  f 

-j-  =  rate  of  evaporation. 

a  —  the  rate  of  evaporation  in  still  air. 

b  —  the  rate  of  increase  with  velocity. 

er  =  the  vapor  pressure  of  the  liquid. 

e  =  the  vapor  pressure  in  the  atmosphere. 

v  =  velocity. 

The  only  difference  between  case  one  and  case  two  is  that  in  case 
one  the  vapor  pressure  of  the  liquid  is  one  of  the  known  or  assumed  factors, 
being  dependent  upon  the  known  temperature  of  the  liquid,  while  in 
case  two,  e}  is  the  vapor  pressure  corresponding  to  the  wet-bulb  tem- 
perature of  the  air. 

This  wet-bulb  or  evaporation  temperature  is  dependent  upon  the  dry- 
bulb  temperature  and  the  moisture  content,  or  upon  the  total  heat  of  the 
air  as  indicated  in  the  previous  paragraph. 

The  effect  of  air  velocity  depends  upon  whether  the  flow  of  air  is 
parallel  to  the  surface  or  perpendicular  to  the  surface  elements.  For  a 
flow  of  air  parallel  to  a  horizontal  surface 

w  =  0.093  (  1  +  ~Q  )  («f  —  e)     (approximately)  (18) 

where 

w  =  pounds  evaporated  per  square  foot  per  hour. 

v  —  velocity  of  atmosphere  over  surfaces,  feet  per  minute. 
e1  =  vapor  pressure  of  the  water  corresponding  to  its  temperature. 

e  =  vapor  pressure  in  the  surrounding  atmosphere. 

For  transverse  flow,  as  across  a  tubular  surface,  the  rate  of  evaporation 
is  nearly  doubled. 

These  relationships  are  indicated  graphically  on  the  chart,  Fig.  3. 

Since  the  difference  in  vapor  pressures  is  substantially  proportional  to 
the  difference  between  the  wet-  and  dry-bulb  temperatures  (i.e.,  the  wet- 
bulb  depression)  the  rate  of  evaporation  is  also,  for  case  two,  substantially 
proportionate  to  the  wet-bulb  depression. 

In  case  two,  the  rate  of  sensible  heat  transfer  from  the  air  to  the  liquid 
to  produce  evaporation  is  substantially  the  same  as  the  rate  of  heat 
transfer  with  the  same  type  of  surface,  without  moisture  being  present, 
but  with  the  same  temperature  differences.  In  other  words,  the  rate  of 
heat  transfer  depends  upon  the  temperature  difference  only,  whether  the 
surface  is  wet  or  not.  For  example,  it  has  been  shown  that  the  rate  of 
heat  transfer  with  air  flowing  across  staggered  coils  (transverse  flow)  may 
be  represented  by  the  formula: 

1 


27 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


where 


heat  transfer  expressed  in  Btu  per  hour  per  square  foot  per  degree 
difference  in  temperature  between  steam  and  air,  for  transverse  flow. 


At  a  velocity  of  400  fpm,  Ut  «  5.8;  at  a  velocity  of  800  fpm,  Ut  =  9.3. 

Referring  to  Fig.  3,  showing  the  rate  of  heat  transmission  by  evapo- 
ration for  different  air  velocities,  it  will  be  noted  that  for  transverse  flow 
there  are  560  Btu  per  hour  per  square  foot  transferred  per  inch  difference 
of  vapor  pressure  at  a  velocity  of  400  fpm,  and  910  Btu  per  hour  per  square 
foot  per  inch  difference  in  vapor  pressure  at  a  velocity  of  800  fpm.  One 
inch  of  vapor  pressure  difference  corresponds  approximately  to  95  deg 
difference  between  the  wet-  and  dry-bulb  temperature.  Dividing  by  95, 


TT1 I11I8I1I1III1 

FIG.  3.    HEAT  TRANSMITTED  BY  EVAPORATION 

the  value  of  5.9  Btu  per  square  foot  per  degree  difference  in  temperature 
is  obtained  for  a  velocity  of  400  fpm,  and  9.55  Btu  per  square  foot  for  a 
velocity  of  800  fpm. 

It  will  be  noted  that  for  these  two  cases  the  heat  transfer  by  evapo- 
ration per  degree  difference  in  temperature  corresponds  almost  exactly 
with  the  heat  transfer  by  convection  coils.  The  similarity  may  be  noted 
by  comparing  the  formula  for  heat  transfer  in  parallel  flow,  where 


0.026 


161 

v 


(20) 


with  the  heat  transfer  by  evaporation  with  parallel  flow.  The  relationship 
will  be  seen  to  be  very  close  in  both  cases  and  would  indicate  that  the  heat 
transfer  by  evaporation  is  actually  brought  about  by  a  process  of  con- 
vection. 

28 


CHAPTER  1 — FUNDAMENTALS  OF  HEATING  AND  AIR  CONDITIONING 

The  difference  in  form  of  the  two  formulae  may  be  due  in  part  to 
errors  in  observation  at  the  higher  and  lower  velocities. 

In  cooling  air  and  condensing  out  the  moisture  therefrom  the  heat 
transfer  is  considerably  more  rapid  than  when  the  air  is  dry  and  no 
moisture  is  condensed.  In  general  the  rate  of  heat  transmission  on  the 
air  side  is  increased  an  amount  which  is  proportionate  to  the  latent  heat 
removed  as  compared  with  the  sensible  heat  removed.  That  is,  if  the 
latent  heat  removed  was  50  per  cent  of  the  sensible  heat  removed,  then 
the  conductivity  of  the  surface  in  contact  with  the  air  would  be  increased 
approximately  50  per  cent. 

REFERENCES 

A  Review  of  Psychrometric  Charts,  by  C.  O.  Mackey  (Heating  and  Ventilating, 
June,  July,  1931). 

A  New  Psychrometric  Chart,  by  C.  A.  Bulkeley  (A.S.H.V.E.  TRANSACTIONS,  Vol.  32, 
1926). 

Air  Conditioning  Applied  to  Cold  Storage  and  a  New  Psychrometric  Chart,  by  C.  A. 
Bulkeley  (Refrigerating  Engineering,  February,  1932). 

Air  Conditioning  Theory,  by  John  A.  Goff  (Refrigerating  Engineering,  January,  1933). 

Rational  Psychrometric  Formulae,  by  W.H.  Carrier  (A.S.M.E.  Transactions,  Vol.  33, 
1911). 

Temperature  of  Evaporation,  by  W.  H.  Carrier  (A.S.H.V.E.  TRANSACTIONS,  Vol.  24, 
1918). 

Principles  of  Engineering  Thermodynamics,  by  Kiefer  and  Stuart. 

Basic  Theory  of  Air  Conditioning,  by  Lawrence  Washington  (Western  Conference  on 
Air  Conditioning,  San  Francisco,  Calif.,  February  9-10,  1933). 

Mixtures  of  Air  and  Water  Vapor,  by  C.  A.  Bulkeley  (Refrigerating  Engineering, 
January,  1933). 

Temperature  of  Evaporation  of  Water  into  Air,  by  W.  H.  Carrier  and  D.  C.  Lindsay 
(A.S.M.E.  Transactions,  1924). 

Chemical  Engineering,  by  Lewis,  Walker  and  McAdams. 

Fan  Engineering,  Buffalo  Forge  Co. 

The  Psychrometric  Chart,  by  E.  V.  Hill  (Aerologist,  April,  May,  June,  1932). 

PROBLEMS  IN  PRACTICE 

1  •  Given  air  at  70  F  dry -bulb  and  50  per  cent  relative  humidity  with  a  baro- 
metric pressure  of  29.00  in.  Hg,  find  the  weight  of  vapor  per  pound  of  dry  air. 

Weight  of  saturated  vapor  per  pound  of  dry  air  =  Wt  =  0.01578  Ib  (Table  5).    Satura- 
tion pressure  of  the  vapor  at  70  F  =  et  =  0.73S6  in.  Hg. 
From  Equation  7r 

0.01578  X  0.5  (29.00  -  0.7386) 

29.00  -  (0.5)  (0.7386) 

W  =  0.00779  Ib  of  vapor  per  pound  of  dry  air  at  70  F  dry-bulb  and  50  per  cent  relative 
humidity. 

Approximate  Method: 

0.01578  X  0.5  =  0.00789  Ib  of  vapor  per  pound  of  dry  air  at  70  F  dry-bulb  and  50  per 

cent  relative  humidity. 

2  •  Given  air  with  a  dry-bulb  temperature  of  80  F,  relative  humidity  of  55  per 
cent,  and  a  barometric  pressure  of  29.92  in.  Hg,  calculate  the  weight  of  a  cubic 
foot  of  the  mixture. 

Weight  of  saturated  vapor  per  cubic  foot  =  0.0015SO  Ib  (Table  5), 

29 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

0.001580  X  0.55  =  0.000869  Ib  =  weight  of  vapor  per  cubic  foot  at  55  per  cent  relative 

humidity. 

Pressure  of  saturated  vapor  at  80  F  =  1.0314  in.  Hg. 

Pressure  of  the  vapor  in  the  mixture  =  1.0314  X  0.55  =  0.567  in.  Hg. 

Pressure  of  the  dry  air  in  the  mixture  =  29.92  -  0.567  ==  29.353  in.  Hg. 

Weight  of  1  cu  ft  of  dry  air  at  80  F  =  -r^r-  =  0.073529  Ib. 

io.oU 

2Q  QCO 

Weight  of  dry  air  in  1  cu  ft  of  the  mixture  =  0.073529  X  i^Sr  =  0.072136  Ib. 

/y.y^ 

0.072136  +  0.000869  =  0.073005  Ib  -  weight  of  1  cu  ft  of  the  mixture. 

3  •  Given  air  with  a  dry-bulb  temperature  of  75  F,  a  relative  humidity  of  60  per 
cent,  and  a  barometric  pressure  of  29-92  in.  Hg,  calculate  the  volume  of  1  Ib 
of  the  mixture. 

Weight  of  saturated  vapor  per  cubic  foot  =  0.001352  Ib  (Table  5). 

0.001352  X  0.6  =  0.0008112  Ib  =  weight  of  vapor  per  cubic  foot  at  60  per  cent  relative 

humidity. 

Pressure  of  saturated  vapor  at  75  F  =  0.8744  in.  Hg. 

Pressure  of  vapor  in  the  mixture  =  0.8744  X  0.6  **  0.525  in.  Hg. 

Pressure  of  dry  air  in  the  mixture  =  29.92  -  0.525  =  29.395  in.  Hg. 

Volume  of  1  Ib  of  dry  air  at  75  F  =  13.48  cu  ft. 

on  QO 
Volume  of  1  Ib  of  dry  air  in  the  mixture  =*  13.48  X  OA  onc.  =  13.72  cu  ft. 

,&y.oyo 

Weight  of  dry  air  in  1  cu  ft  of  the  mixture  =    lV,0    =  0.072886  Ib. 

Lo,t  A 

0.072886  +  0.000811  «  0.073697  Ib  »  weight  of  1  cu  ft  of  the  mixture. 

A  A»7oafV7  ~  13.57  cu  ft  =  volume  of  1  Ib  of  the  mixture. 
i/.u/ooy/ 

Approximate  Method: 

Volume  of  1  Ib  of  saturated  air  at  75  F  «  13.88  cu  ft. 

Volume  of  1  Ib  of  dry  air  at  75  F  =  13.48  cu  ft. 

Difference  in  volume  =    0.40  cu  ft. 

Relative  humidity  -  60  per  cent. 
0.40  X  0.6  =  0.24  cu  ft. 

13.48  +  0.24  =  13.72  cu  ft  «  volume  of  1  Ib  of  the  mixture. 
The  degree  of  approximation  is  evident. 

4  •  Given  saturated  air  at  a  temperature  of  75  F  and  a  barometric  pressure  of 
29.92  in.  Hg,  determine  the  total  heat  of  the  mixture  per  pound  of  dry  air. 

From  Equation  11  and  Table  5, 

Cpa  =  mean  specific  heat  at  constant  pressure  of  dry  air  =  0.24. 
.    feg  =  latent  heat  of  vaporization  at  the  wet-bulb  temperature  ==  1050.1  Btu  per  Ib. 
W<L   =  weight  of  water  vapor  mixed  with  each  pound  of  dry  air  =  0.01877  Ib. 
2  =  0.24  (75  -  0)  +  (0.01877)  (1050.1). 
S  =  37.71  Btu  per  Ib  of  dry  air. 

5  •  Given  ah*  at  85  F  dry-bulb  temperature,  75  F  wet-bulb  temperature,  and  a 
barometric  pressure  of  29.92  in.  Hg;  determine  the  total  heat  of  the  mixture 
per  pound  of  dry  air. 

From  Equation  10  and  Table  5, 
CPa  =  0.24. 
Affg  =  1050.1  Btu. 

30 


CHAPTER  1 — FUNDAMENTALS  OF  HEATING  AND  AIR  CONDITIONING 

Relative  humidity  =  62.3  per  cent  (from  psychro metric  chart). 
W  =  0.02634  X  0.623  =  0.01641  grains  of  moisture  per  Ib  of  dry  air. 
S  =  0.24  (85  -  0)  -f  0.01641  [1050.1  -f  0.45  fS5  -  75)]. 
2  =  37.71  Btu  per  pound  of  dry  air. 

It  will  be  seen  from  Questions  4  and  5  that  the  total  heat  content  is  a  function  of  the 
wet-bulb  temperature. 

6  •  It  is  desired  to  maintain  a  temperature  of  80  F  and  a  relative  humidity  of 
50  per  cent  in  a  factory  where  the  equipment  gives  off  6,000  Btu  per  hour.     If 
the  entering  air  is  at  70  F,  determine  the  relative  humidity,  and  the  pounds  of 
air  required  per  hour. 

Air  at  80  F  and  50  per  cent  relative  humidity  contains  77  grains  of  moisture  per  pound. 
At  70  F  and  77  grains  of  moisture  per  pound,  the  relative  humidity  is  70  per  cent. 

Total  heat  above  zero  in  the  mixture  at  80  F  and  50  per  cent  relative  humidity  =  31.2 
Btu  per  pound. 

Total  heat  above  zero  in  the  mixture  at  70  F  and  70  per  cent  relative  humidity  =  28.8 
Btu  per  pound. 

31.2  -  28.8  =  2.4  Btu  to  be  removed  per  pound  of  air. 
6000  Btu  =  heat  given  off  by  equipment  per  hour. 

6000 

=  2500  Ib  of  air  required  per  hour. 

£A 

7  •  From  the  data  given  in  Question  6,  calculate  the  approximate  cubic  feet 
of  air  required  per  minute. 

Volume  of  1  Ib  of  saturated  air  at  70  F  =  13.69  cu  ft  (Table  5) 
Volume  of  1  Ib  of  dry  air  at  70  F  =  13.35  cu  ft. 

Difference  in  volume  =    0.34  cu  ft. 

Relative  humidity  =  70  per  cent. 
0.34  X  0.7  =  0.24  cu  ft. 

13.35  +  0.24  =  13.59  cu  ft,  volume  of  1  Ib  of  mixture  at  70  F  and  70  per  cent  relative 
humidity  (approximate). 

From  Question  6  the  air  required  per  hour  =  2500  Ib. 
2500  X  13.59 


60 


566.25  cu  ft  per  minute  required. 


8  •  Given  1  Ib  of  dry  air  at  78  F  and  a  barometric  pressure  of  29.92  in.  Hg; 
calculate  the  volume.  If  the  temperature  is  raised  to  96  F  and  the  volume 
remains  constant,  what  will  be  the  new  pressure,  P2,  in  in.  Hg? 

PV  =  WRT. 

R  (for  air)  =  53.34. 

W  =  1  Ib. 

P  —  absolute  pressure,  pounds  per  square  foot. 

„  _  1  X  53.34  X  (78  +  460) 
29.92  X  0.491  X  144 

V  =  13.57  cu  ft  =  volume  of  1  Ib. 


PI     rr     z      TI 

(96  +  460)  (29.92  X  0.491  X  144) 
2  (78  -f  460)  (0.491  X  144) 

P2  »  30.90  in.  Hg.  - 

31 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

9  •  Given  saturated  air  at  a  temperature  of  75  F  and  a  barometric  pressure  of 
29.92  in.  Hg;  determine  the  heat  content  of  the  mixture  per  pound  of  dry  air, 
including  the  heat  content  of  the  liquid  above  32  F. 

From  Equation  12, 

/;  =  0.24  (/  -  0)  +  W  (1059.2  -f  0.450- 
where 

hs  =  1059.2  -f  0.45/  (Empirical  equation  derived  from  Keenan's  Steam  Tables.) 

/  =  75  F. 

II'  =  0.01S77  Ib  of  water  vapor  (Table  5). 

h  -  0.24  (75  -  0)  -|-  0.01877  (1059.2  4-  0.45  X  75). 

h  =  38.51  Btu  per  pound  of  dry  air. 


Chapter  2 

VENTILATION 
AND  AIR  CONDITIONING  STANDARDS 

litiation  of  Air,  Heat  Regulation  in  Man,  Effects  of  Heat, 
Effects  of  Cold,  Temperature  Changes,  Acclimatization, 
W'armth  and  Comfort,  Effective  Temperature,  Comfort  Chart, 
Comfort  Line,  Comfort  Zone,  Application  of  Comfort  Chart, 
A.S.H.V.E.  Ventilation  Standards,  Natural  and  Mechanical 
Ventilation,  Recirculation,  Ultra-Violet  Radiation  and  lonisa- 
tion,  Heat  and  Moisture  Losses 

VENTILATION  is  defined  in  part  as  "the  process  of  supplying  or 
removing  air  by  natural  or  mechanical  means  to  or  from  any  space." 
(See  Chapter  41.)  The  word  in  itself  implies  quantity  but  not  necessarily 
quality.  From  the  standpoint  of  comfort  and  health,  however,  the 
problem  is  now  considered  to  be  one  of  securing  air  of  the  proper  quality 
rather  than  of  supplying  a  given  quantity. 

The  term  air  conditioning  in  its  broadest  sense  implies  control  of  any  or 
all  of  the  physical  or  chemical  qualities  of  the  air.  More  particularly,  it 
includes  the  simultaneous  control  of  temperature,  humidity,  movement, 
and  purity  of  the  air.  The  term  is  broad  enough  to  embrace  whatever 
other  additional  factors  may  be  found  desirable  for  maintaining  the 
atmosphere  of  occupied  spaces  at  a  condition  best  suited  to  the  physio- 
logical requirements  of  the  human  body. 

VITIATION  OF  AIR 

Under  the  artificial  conditions  of  indoor  life,  the  air  undergoes  certain 
physical  and  chemical  changes  which  are  brought  about  by  the  occupants 
themselves.  The  oxygen  content  is  somewhat  reduced,  and  the  carbon 
dioxide  slightly  increased  by  the  respiratory  processes.  Organic  matter, 
which  is  usually  perceived  as  odors,  comes  from  the  nose,  mouth,  skin 
and  clothing.  The  temperature  of  the  air  is  increased  by  the  metabolic 
processes,  and  the  humidity  raised  by  the  moisture  emitted  from  the  skin 
and  lungs.  Moreover,  according  to  latest  researches1,  there  is  a  marked 
decrease  in  both  positive  and  negative  ions  in  the  air  of  occupied  rooms. 

Contrary  to  old  theories,  the  usual  changes  in  oxygen  and  carbon 
dioxide  are  of  no  physiological  concern  because  they  are  much  too  small 
even  under  the  worst  conditions.  The  amount  of  carbon  dioxide  in  air  is 
often  used  in  ventilation  work  as  an  index  of  odors  of  human  origin,  but 


*See  A.S.H.V.E.  research  paper  entitled  Changes  in  Ionic  Content  in  Occupied  Rooms  Ventilated  by 
Natural  and  Mechanical  Methods,  by  C.  P.  Yaglou,  L.  C.  Benjamin  and  S.  P.  Choate  (A.S.H.V.E,  TRANS- 
ACTIONS, Vol.  37,  1931). 

33 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

the  information  it  affords  rarely  justifies  the  labor  involved  in  making  the 
observation2.  Little  is  known  of  the  identity  and  physiological  effects  of 
the  organic  matter  given  off  in  the  process  of  respiration.  The  former 
belief  that  the  discomfort  experienced  in  confined  spaces  was  due  to  some 
toxic  volatile  matter  in  the  expired  air  is  now  limited,  in  the  light  of 
numerous  researches,  to  the  much  less  dogmatic  view  that  the  presence  of 
such  a  substance  has  not  been  demonstrated.  The  only  certain  fact  is 
that  expired  and  transpired  air  is  odorous  and  offensive,  and  it  is  capable 
of  producing  loss  of  appetite  and  a  disinclination  for  physical  activity. 
These  reasons  alone,  whether  aesthetic  or  physiological,  are  sufficient  to 
warrant  a  desire  for  proper  air  conditions. 

A  certain  part  of  the  dissemination  of  disease  which  occurs  in  confined 
spaces  is  caused  by  the  emission  of  pathogenic  bacteria  from  infected 
persons.  Infections  by  droplets  from  coughing  and  sneezing  constitute  a 
limited  mode  of  transmission  in  the  immediate  vicinity  of  the  infected 
person.  Experiments  have  shown  that  the  mouth  spray  is  a  coarse  rain 
which  settles  down  quickly.  The  contamination  is  local  and  the  problem 
is  considered  to  be  largely  one  of  contact  infection  rather  than  air-borne 
infection. 

The  primary  factors  in  air  conditioning  work,  in  ^the  absence  of  any 
specific  contaminating  source,  are  temperature,  humidity,  air  movement 
and  body  odors.  As  compared  with  these  physical  factors,  the  chemical 
factors  are,  as  a  general  rule,  of  secondary  importance. 

HEAT  REGULATION  IN  MAN 

The  importance  of  temperature,  humidity  and  air  movement  arises 
from  the  profound  influence  which  these  factors  exert  upon  body  tem- 
perature, comfort  and  health.  Body  temperature  is  a  resultant  of  the 
balancing  action  between  its  heat  production  and  its  heat  loss.  ^  The  heat 
resulting  from  the  combustion  of  food  within  the  body  maintains  its 
temperature  well  above  that  of  the  surrounding  air.  At  the  same  time, 
heat  is  constantly  lost  from  the  body  by  radiation,  conduction  and 
evaporation.  Since,  under  ordinary  conditions,  the  body  temperature  is 
maintained  at  its  normal  level  of  about  98.6  F,  the  heat  production  must 
be  balanced  by  the  heat  loss.  In  healthy  persons  this  takes  place  auto- 
matically by  the  action  of  the  heat  regulating  mechanism. 

According  to  the  general  view,  special  areas  in  the  skin  are  sensitive  to 
temperature.  Nerve  courses  carry  the  sense  impressions  to  the  brain  and 
the  response  comes  back  over  another  set  of  nerves,  the  motor  nerves,  to 
the  musculature  and  to  all  the  active  tissues  in  the  body,  including  the 
endocrine  glands.  In  this  way,  a  two-sided  mechanism  controls  the  body 
temperature  by  (1)  regulation  of  internal  heat  production  (chemical 
regulation),  and  (2)  regulation  of  heat  loss  by  means  of  automatic  varia- 
tion in  the  rate  of  cutaneous  circulation  and  the  operation  of  the  sweat 
glands  (physical  regulation).  The  mechanisms  of  adjustment  are  complex 
and  little  understood  at  the  present  time.  Coordination  of  these  dif- 
ferent mechanisms  seems  to  vary  greatly  with  different  air  conditions. 


'Indices  of  Air  Change  and  Air  Distribution,  by  F.  C.  Houghten  and  J.  L.  Blackshaw  (A.S.H.V.E. 
Journal  Section,  Heating,  Piping  and  Air  Conditioning,  June,  1933,  p.  324). 


CHAPTER  2 — VENTILATION  AND  AIR  CONDITIONING  STANDARDS 

With  rising  air  temperatures  up  to  75  F  or  80  F,  metabolism,  or  internal 
heat  production,  is  decreased3,  probably  by  an  inhibitory7  action  on  heat 
producing  organs,  especially  the  adrenal  glands,  which  seem  to  exert  the 
major  influence  on  basic  combustion  processes  in  the  body.  The  blood 
capillaries  in  the  skin  become  dilated  by  reflex  action  of  the  vasomotor 
nerves,  allowing  more  blood  to  flow  into  the  skin,  and  thus  increase  its 
temperature  and  consequently  its  heat  loss.  The  increase  in  peripheral 
circulation  is  at  the  expense  of  the  internal  organs.  If  this  method  of 
cooling  is  not  in  itself  sufficient,  the  stimulus  is  extended  to  the  sweat 
glands  which  allow  water  to  pass  through  the  surface  of  the  skin,  where  it 
is  evaporated.  This  method  of  cooling  is  the  most  effective  of  all,  as  long 
as  the  humidity  of  the  air  is  sufficiently  low  to  allow  for  evaporation.  In 
high  humidities,  where  the  difference  between  the  dew-point  temperature 
of  the  air  and  body  temperature  is  not  sufficient  to  allow  rapid  evapora- 
tion, equally  good  results  may  be  obtained  by  increasing  the  air  move- 
ment, and  hence  the  heat  loss  by  conduction  and  evaporation. 

In  cold  environments,  in  order  to  keep  the  body  warm  there  is  an  actual 
increase  in  metabolism  brought  about  partly  by  voluntary  muscular  con- 
tractions (shivering)  and  partly  by  an  involuntary  reflex  upon  the  heat 
producing  organs.  The  surface  blood  vessels  become  constricted,  and 
the  blood  supply  to  the  skin  is  curtailed  by  vasomotor  shifts  to  the  internal 
organs  in  order  to  conserve  body  heat. 

EFFECTS  OF  HEAT 

Although  the  human  organism  is  capable  of  adapting  itself  to  variations 
in  environmental  conditions,  its  ability  to  maintain  heat  equilibrium  is 
limited.  The  heat  regulating  center  fails,  for  instance,  if  the  external 
temperature  is  so  abnormally  high  that  bodily  heat  cannot  be  eliminated 
as  fast  as  it  is  produced.  Part  of  it  is  retained  in  the  body,  causing  a  rise 
in  skin  and  deep  tissue  temperature,  an  increase  in  the  heart  rate,  and 
accelerated  respiration.  (See  Table  1.)  In  extreme  conditions,  the 
metabolic  rate  is  markedly  increased  owing  to  the  excessive  rise  in  body 
temperature4,  and  a  vicious  cycle  results  which  may  eventually  lead  to 
serious  physiologic  damage. 

Examples  of  this  are  met  with  in  unusually  hot  summer  weather  and  in 
hot  industries  where  the  radiant  heat  from  hot  objects  renders  heat  loss 
from  the  body  by  radiation  and  convection  impossible.  Consequently, 
the  workers  depend  entirely  on  evaporation  for  the  elimination  of  body 
heat.  They  stream  with  perspiration  and  drink  liquids  abundantly  to 
replace  the  loss. 

One  of  the  most  deleterious  effects  of  high  temperatures  is  that  the 
blood  is  diverted  from  the  internal  organs  to  the  surface  capillaries,  in 
order  to  serve  in  the  process  of  cooling.  This  affects  the  stomach,  heart, 
lungs  and  other  vital  organs,  and  it  is  believed  that  the  feeling  of  lassitude 
and  discomfort  experienced  is  due  to  the  anaemic  condition  of  the  brain. 


*Heat  and  Moisture  Losses  From  the  Human  Body  and  Their  Relation  to  Air  Conditioning  Problems* 
by  F.  C.  Houghten,  W,  W.  Teague,  W.  E.  Miller,  and  W.  P.  Yant  (A.S.H.V.E,  TRANSACTIONS,  Vol.  35, 
1929,  p.  245). 

*ThennaI  Exchanges  Between  the  Human  Body  and  Its  Atmospheric  Environment,  by  F.  C.  Houghteiu 
W.  W.  Teague,  W.  E.  MUfer,  and  W.  P.  Yant  {The  American  Journal  of  PJfcyswrfagy,  Vol.  8&,  No,  £,  April. 
1929). 

35 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


TABLE  1.    PHYSIOLOGICAL  RESPONSES  TO  HEAT  OF  MEN  AT  REST  AND  AT  \VoRKa 

I  '  j  MEN  AT  WORK 


ACTUAL 

MEN  A.T  RE 

ST 

90,( 

XX)  FT-LB  OF  V 

ORE  PER  HOU 

R 

EFFECTIVE 
TEMP. 

CHEEK 
TEMP 
(DEG 
FAHR) 

Rise  in 
Rectal 
Temp 
(Deg 
Fahrper 
Hour) 
- 

Increase 
in  Pulse 
Rate 
(Beats  per 
Mb  per 
Hour) 

Approximate 
Loss  in  Body 
Weight  by 
Perspiration 
(Lb  perHr) 

Total  Work 
Accomplished 

(Ft-lb) 

Rise  in 
Body  Temp 
(Deg  Fahr 
per  Hr) 

Increase  in 
Pulse  Rate 
(Beats  per 
Min  per  Hr) 

Approximate 
Loss  in  Body 
Wt.  by  Per- 
spiration (Lb 
per  Hr} 

60 

225,000 

0.0 

6 

0.5 

70 

0.0 

0 

0.2 

225,000 

0.1 

7 

0.6 

80 

96.7 

0.0 

0 

0.3 

209,000 

0.3 

11 

0.8 

85 

96.6 

0.1 

1 

0.4 

190,000 

0.6 

17 

1.1 

90 

97.0 

0.3 

4 

0.5 

153,000 

1.2 

31 

1.5 

95 

97.6 

0.9 

15 

0.9 

102,000 

2.3 

61 

2.0 

100 

99.6 

2.2 

40 

1.7 

67,000 

4.0 

103b 

2.7 

105 

104.7 

4.0 

83 

2.7 

49,000 

6.  Ob 

158^ 

3.5b 

110 



5.9t» 

137^ 

4.  0^ 

37,000 

8.5b 

237° 

4.4*> 

«Data  by  A.S.H.V.E.  Research  Laboratory. 

bComputed  va^e  from  exposures  lasting  less  than  one  hour. 

The  stomach  loses  some  of  its  power  to  act  upon  the  food,  owing  to  a 
diminished  secretion  of  gastric  juice,  and  there  is  a  corresponding  loss  in 
the  antiseptic  and  antifermentive  action  which  favors  the  growth  of 
bacteria  in  the  intestinal  tract5.  These  are  considered  to  be  the  potent 
factors  in  the  increased  susceptibility  to  gastro-intestinal  disorders  in  hot 
summer  weather.  The  vie  fim  may  lose  appetite  and  suffer  from  indiges- 
tion, headache  and  general  enervation,  which  may  eventually  lead  to  a 
premature  old  age. 

In  warm  atmospheres,  particularly  during  physical  work,  a  considerable 
amount  of  chloride  is  lost  from  the  system  through  sweating.  The  loss  of 
this  substance  may  lead  to  attacks  of  cramps,  unless  the  salts  are  replaced 
in  the  drinking  water.  In  order  to  relieve  both  cramps  and  fatigue, 
Moss6  recommends  the  addition  of  6  grams  of  sodium  chloride  and  4  grams 
of  potassium  chloride  to  a  gallon  of  water. 

The  deleterious  physiologic  effects  of  high  temperatures  exert  a  power- 
ful influence  upon  physical  activity,  accidents,  sickness  and  mortality. 
Both  laboratory  and  field  data  show  clearly  that  physical  work  in  warm 
atmospheres  is  a  great  effort,  and  that  production  falls  progressively  as 
the  temperature  rises.  The  incidence  of  industrial  accidents  reaches  a 
minimum  at  about  68  F,  increasing  above  and  below  that  temperature. 
Sickness  and  mortality  rates  increase  progressively  as  the  temperature 


rises. 


EFFECTS  OF  COLD 


The  action  of  cold  on  human  beings  is  not  well  known.  Cold  affects  the 
human  organism  in  two  ways:  (1)  through  its  action  on  the  body  as  a 
whole,  and  (2)  through  its  action  on  the  mucous  membranes  of  the  upper 
respiratory  tract.  Little  exact  information  is  available  on  the  latter. 

On  exposure  to  cold,  the  loss  of  heat  is  increased  considerably  and  only 


^Influence  of  Effective  Temperature  upon  Bactericidal  Action  of  Gasto-Intestinal  Tract,  by  Arnold  and 
Brody  (Proceedings  Society  Exp.  Biol.  Med.  Vol.  24,  1927,  p.  832). 

6Some  Effects  of  High  Air  Temperatures  upon  the"  Miner,  by  K.'N.  Moss  (Transactions  institute  of 
Mining  Engineers,  Vol.  66,  1924,  p.  284). 

36 


CHAPTER  2 — VENTILATION  AND  AIR  CONDITIONING  STANDARDS 

within  certain  limits  is  compensation  possible  by  increased  heat  produc- 
tion and  decreased  peripheral  circulation.  The  rectal  temperature  often 
rises  upon  exposure  to  cold  but  the  pulse  rate  and  skin  temperature  fall. 
The  blood  pressure  increases,  owing  to  constriction  in  the  peripheral 
vessels  and  to  thickening  of  the  blood.  The  subcutaneous  tissues  and 
muscles  form  reservoirs  for  storing  the  water  which  leaves  the  blood.  In 
extremely  cold  atmospheres  compensation  becomes  inadequate.  The 
body  temperature  falls  and  the  reflex  irritability  of  the  spinal  cord  is 
markedly  affected.  The  organism  may  finally  pass  into  an  unconscious 
state  which  ends  in  death. 

Cannon  showed  that  excessive  loss  of  heat  is  associated  with  increased 
activity  of  the  adrenal  medulla7.  The  extra  output  of  adrenin  hastens 
heat  production  which  protects  the  organism  against  cooling.  Bast8 
found  a  degeneration  of  thyroid  and  adrenal  glands  upon  exposure  to  cold. 

Effects  of  Temperature  Changes 

A  moderate  amount  of  variability  in  temperature  is  known  to  be 
beneficial  to  health,  comfort,  and  the  performance  of  physical  and  mental 
work.  On  the  other  hand,  extreme  changes  in  temperature,  such  as  those 
experienced  in  passing  from  a  warm  room  to  the  cold  air  out  of  doors, 
appear  to  be  harmful  to  the  tissues  of  the  nose  and  throat  which  are  the 
portals  for  the  entry  of  respiratory  diseases. 

Experiments  show  that  chilling  causes  a  constriction  of  the  blood 
vessels  of  the  palate,  tonsils  and  throat,  which  is  accompanied  by  a  fall 
in  the  temperature  of  the  tissues.  On  rewarming,  the  palate  and  throat 
do  not  always  regain  their  normal  temperature  and  blood  supply.  This 
anaemic  condition  favors  bacterial  activity  and  it  is  believed  to  play  a 
part  in  the  inception  of  the  common  cold  and  other  respiratory  diseases. 
It  is  believed  that  the  lowered  resistance  is  due  to  a  diminution  in  the 
number  and  phagocytic  activity  ,of  the  leucocytes  (white  blood  cells) 
brought  about  by  exposure  to  cold  and  by  changes  in  temperature. 

Sickness  records  in  industries  seem  to  strengthen  this  belief.  The 
Industrial  Fatigue  Research  Board  of  England9  found  that  in  workers 
exposed  to  high  temperatures  and  to  changes  in  temperature,  namely, 
steel  melters,  puddlers,  and  tin-plate  rnillmen,  there  is  an  excess  of  all 
sickness,  the  excess  among  the  puddlers  being  due  chiefly  to  respiratory 
diseases  and  rheumatism.  The  causative  factor  was  not  the  heat  itself 
but  the  sudden  changes  in  temperature  to  which  the  workers  were  exposed. 
The  tin-plate  millmen  who  were  not  exposed  to  chills,  since  they  work 
almost  continuously  throughout  the  shift,  had  no  excess  of  rheumatism 
and  respiratory  diseases.  On  the  other  hand,  the  blast-furnacemen,  who 
work  mostly  in  the  open,  showed  more  respiratory  sickness  than  the  steel 
workers.  This  experience  in  British  factories  is  well  in  accord  with  the 
findings  in  American  industries10.  According  to  these  data  the  highest 


^Studies  on  the  Condition  of  Activity  of  Endocrine  Glands,  by  W.  B.  Cannon,  A.  Guerido,!  S.  W.  Britton 
and  E.  M.  Bright  (American  Journal  of  Physiology,  Vol.  79,  1926,  p.  466). 

8St tidies  in  Exhaustion  Due  to  Lack  of  Sleep,  by  T.  H.  Bast,  J.  S.  Supernaw,  B.  Lieberman  and  J.  Munro 
(American  Journal  of  Physiology,  Vol.  85,  1928,  p,  135). 

9Fatigue  and  Efficiency  in  the  Iron  and  Steel  Industry,  by  H.  M.  Veraon  {Industrial  Fatigue  Research 
Board,  Report  No.  5,  1920,  London). 

MIron  Foundry  Workers  Show  Highest  Percentage  of  Deaths  from  Pneumonia  {Statistical  Bulletin, 
Metropolitan  Life  Insurance  Company,  1928). 

37 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

pneumonia  death  rate  is  associated  with  dust,  extreme  heat,  exposure  to 
cold,  and  to  sudden  changes  in  temperature. 

ACCLIMATIZATION 

Acclimatization  and  the  factor  of  psychology  are  two  important  in- 
fluences in  air  conditioning  which  cannot  be  ignored.  The  first  is  man's 
ability  to  adapt  himself  to  changes  in  air  conditions;  the  second  is  an 
intangible  matter  of  habit  and  suggestion. 

Some  persons  regard  the  unnecessary  endurance  of  cold  as  a  virtue. 
They  believe  that  the  human  organism  can  adapt  itself  to  a  wide  range  of 
air  conditions  with  no  apparent  discomfort  or  injury  to  health.  In  the 
light  of  the  present  knowledge  of  air  conditioning  these  views  are  not 
justified.  Acclimatization  to  extreme  conditions  involves  a  strain  upon 
the  heat  regulating  system  and  it  interferes  with  the  normal  physiologic 
functions  of  the  human  body.  Thousands  of  years  in  the  heat  of  Africa 
do  not  seem  to  have  acclimatized  the  Negro  to  a  temperature  averaging 
80  F.  The  same  holds  true  of  northern  races  with  respect  to  cold,  although 
the  effects  are  mitigated  by  artificial  control.  All  this  seems  to  indicate 
that  adaptation  to  a  climate  averaging  between  60  and  80  F  is  a  very 
primitive  trait11. 

Within  these  limits,  however,  there  does  occur  a  definite  adaptation  to 
external  temperature  level.  People  and  animals  raised  under  conditions 
of  tropical  moist  heat  have  a  lower  rate  of  heat  production  than  do  those 
who  grow  up  in  cooler  environments.  This  causes  them  to  stand  chilling 
poorly  as  they  are  unable  to  quickly  increase  internal  combustion  to  keep 
up  the  body  temperature.  For  this  reason  they  have  trouble  standing 
the  cold,  stormy  weather  of  the  temperate  zones,  and  when  exposed  to  it 
are  very  susceptible  to  respiratory  infections.  Likewise,  people  living  in 
cool  climates  suffer  greatly  in  the  moist  heat  of  the  tropics  until  their 
adrenal  activity  has  slowed  down.  Within  a  couple  of  years,  however, 
they  find  themselves  standing  the  heat  much  better  and  disliking  cold. 
They  become  acclimated  by  a  definite  change  in  the  combustion  level 
within  the  body12. 

In  certain  individuals  the  psychologic  factor  is  more  powerful  than 
acclimatization.  A  fresh  air  fiend  may  suffer  in  3.  room  with  windows 
closed  regardless  of  the  quality  of  the  air.  As  a  matter  of  fact,  instances 
are  known  in  which  paid  subjects  refused  to  stay  in  a  windowless  but 
properly  conditioned  experimental  chamber  because  the  atmosphere  felt 
suffocating  to  them  upon  entering  the  room. 

WARMTH  AND  COMFORT 

The  temperature,  humidity,  and  motion  of  the  air,  and  the  radiation 
between  a  person  and  surrounding  hot  or  cold  surfaces,  taken  together, 
determine  his  feeling  of  warmth  and  influence  his  elimination  of  body 
heat.  In  other  words,  the  temperature  sensations  of  the  human  body 
depend  not  only  on  the  temperature  of  the  surrounding  air  as  registered 
by  a  dry-bulb  thermometer,  but  also  upon  the  temperature  indicated  by 

"Civilization  and  Climate,  by  Ellsworth  Huntington,  Yale  University  Press,  1924. 
"Air  Conditioning  it  its  Relation  to  Human  Welfare,  by  C.  A.  Mills,  M.D.  (A.S.H.V.E.  Journal  Section, 
Heating,  Piping  and  Air  Conditioning,  April,  1934). 

38 


CHAPTER  2 — VENTILATION  AND  AIR  CONDITIONING  STANDARDS 

a  wet-bulb  thermometer.  Dry  air  at  a  relatively  high  temperature  may 
feel  cooler  than  air  of  considerably  lower  temperature  with  a  high  mois- 
ture content.  Air  motion  makes  any  moderate  condition  feel  cooler. 

On  the  other  hand,  in  cold  environments  an  increase  in  humidity 
produces  a  cooler  sensation.  The  dividing  line  at  which  humidity  has  no 
effect  upon  comfort  varies  with  the  air  velocity  and  is  about  46  F  (dry- 
bulb)  for  still  air  and  about  51,  56  and  59  F  for  air  velocities  of  100,  300 
and  500  fpm,  respectively. 

Thermo- Equivalent  Conditions 

Combinations  of  temperature,  humidity  and  air  movement  which  pro- 
duce the  same  feeling  of  warmth  are  called  thermo-equivalent  con- 
ditions. A  series  of  tests13'  14*  15  has  been  carried  out  in  the  psychrometric 
rooms  of  the  A.S.H.V.E.  Research  Laboratory,  Pittsburgh,  in  order  to 
determine  the  equivalent  conditions  met  with  in  general  air  conditioning 
work.  These  show  that  this  newly-developed  scale  of  thermo-equivalent 
conditions  not  only  indicates  the  sensation  of  warmth,  but  also  determines 
the  physiological  effects  on  the  body  induced  by  heat  and  cold.  "  For  this 
reason,  it  is  called  the  effective  temperature  scale  or  index. 

Effective  temperature  is  an  index  of  warmth  or  cold.  It  is  not  in  itself 
an  index  of  comfort,  as  it  is  often  assumed  to  be,  nor  are  the  effective  tem- 
perature lines  necessarily  lines  of  equal  comfort.  This  is  true  because,  in 
determining  this  index,  the  subjects  compared  not  the  relative  comfort, 
but  rather  the  relative  warmth  or  cold  of  various  air  conditions.  Moist 
air  at  a  comparatively  low  temperature,  and  dry  air  at  a  higher  tempera- 
ture may  each  feel  as  warm  as  air  of  an  intermediate  temperature  and 
humidity,  but  the  comfort  experienced  in  the  three  air  conditions  would  be 
different,  although  the  effective  temperature  is  the  same. 

Under  extreme  humidity  conditions  there  seems  to  be  a  difference  be- 
tween sensations  of  absolute  comfort  and  of  the  proper  degree  of  warmth. 
In  other  words,  human  beings  are  not  necessarily  comfortable  when  the 
air  is  neither  too  warm  nor  too  cold.  Air  of  proper  warmth  may,  for  in- 
stance, contain  excessive  water  vapor,  and  in  this  way  interfere  with  the 
normal  physiologic  loss  of  moisture  from  the  skin,  leading  to  damp  skin 
and  clothing  and  producing  more  or  less  discomfort;  or  the  air  may  be 
excessively  dry,  producing  appreciable  discomfort  to  the  mucous  mem- 
brane of  the  nose  and  to  the  skin  which  dries  up  and  becomes  chapped 
from  too  rapid  loss  of  moisture.  According  to  the  comfort  experiments 
first  conducted  at  the  A.S.H.V.E.  Laboratory16  in  the  U.  S.  Bureau  of 
Mines,  Pittsburgh,  and  later  studies  at  the  Harvard  School  of  Public 
Health17  in  Boston,  effective  temperature  appears  to  be  a  fair  index  of 
comfort  also,  particularly  within  a  humidity  range  of  30  to  60  per  cent, 
approximately. 

"Determining  Lines  of  Equal  Comfort,  by  F.  C.  Houghteu  and  C.  P.  Yagloglou  (A.S,H,V.E.  TRANS- 
ACTIONS, Vol.  29,  1923,  p.  361). 

"Cooling  Effect  on  Human  Beings  by  Various  Air  Velocities,  by  F.  C.  Houghten  and  C.  P.  Yaglogtou 
(A.SwH.V.E.  TRANSACTIONS,  Vol.  30,  1924,  p.  193), 

"Effective  Temperature  with  Clothing,  by  C.  P.  Yagloglou  and  W.  E.  Miller  (A.S.H.V.E.  TRANS- 
ACTIONS, Vol.  31,  1925,  p.  89). 

^Determination  of  the  Comfort  Zone  With  Further  Verification  of  Effective  Temperatures  Within  This 
Zone,  by  F.  C.  Houghten  and  C.  P.  Yaglogiou  (A-S.H.V.E.  TRANSACTIONS,  Vol.  29,  1923,  p.  361). 

*rrhe  Summer  Comfort  Zone;  Climate  and  Clothing,  by  C,  P.  Yagloa  and  Philip  Drinker 
(A.S.H.V.E.  TRANSACTIONS,  Vot  35, 1959). 

39 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


dJ 

-aoo: 


9 
.705 


FIG.  1.    THERMOMETRIC  OR  EFFECTIVE  TEMPERATURE  CHART  SHOWING  NORMAL  SCALE 
.       OF  EFFECTIVE  TEMPERATURE.    APPLICABLE  TO  INHABITANTS  OF  THE 
UNITED  STATES  UNDER  FOLLOWING  CONDITIONS: 

A.   Clothing:   Customary  indoor  clothing.   B.   Activity:   Sedentary  or  light  muscular  work.    C.  Heating 
Methods:    Convection  type,  «.«.,  warm  air,  direct  steam  or  hot  water  radiators,  plenum  systems. 

40 


CHAPTER  2 — VENTILATION  AND  AIR  CONDITIONING  STANDARDS 

Definition  of  Effective  Temperature 

Briefly,  effective  temperature  may  be  defined  as  an  arbitrary  index  of  the 
degree  of  warmth  or  cold  felt  by  the  human  body  in  response  to  tempera- 
ture, humidity,  and  movement  of  the  air.  Effective  temperature  is  not  a 
true  temperature  of  the  air  but  an  index  which  combines  temperature, 
humidity  and  air  motion  in  a  single  value.  The  numerical  value  of  the 
effective  temperature  index  for  any  given  air  condition  is  fixed  by  the 
temperature  of  saturated  air  which,  at  a  velocity  or  turbulence  of  15  to 
25  fpm,  induces  a  sensation  of  warmth  or  cold  like  that  of  the  given 
condition.  -  Thus,  any  air  condition  has  an  effective  temperature  of 
65  deg  when  it  induces  a  sensation  of  warmth  like  that  experienced  in 
practically  still  air  at  65  F  saturated  with  moisture. 

In  all  reports  of  the  A.S.H.V.E.  Research  Laboratory,  the  term  still  air 
signifies  the  minimum  air  movement  it  was  possible  to  obtain  in  the 
Laboratory's  psychrometric  chamber.  Actually,  the  air  motion  was 
between  15  and  25  fpm  in  all  experiments,  without  qualification,  as 
measured  by  the  Kata  thermometer.  This  was  not  a  linear  movement  of 
air  but  it  represented  the  turbulence  or  eddy  currents  produced  by  the  air 
change.  Even  in  tightly  sealed  rooms,  the  natural  air  movement  is  not 
likely  to  fall  below  10  fpm  so  long  as  there  is  a  temperature  or  pressure 
difference  between  the  air  inside  and  that  outside  the  room. 

Fig.  1  shows  the  results  obtained  at  the  A.S.H.V.E.  Research  Labora- 
tory in  a  single  chart,  the  so-called  thermometric  chart.  The  equivalent 
conditions  or  effective  temperature  lines  are  shown  by  the  short  cross- 
lines.  The  difference  between  the  effective  temperature  for  still  air  and 
for  moving  air,  of  any  velocity,  represents  the  cooling  resulting  from  that 
air  velocity.  This  thermometric  chart  applies  to  average  normal  and 
healthy  persons  adapted  to  American  living  and  working  conditions.  It 
is  limited  to  sedentary  or  light  muscular  activity,  and  to  rooms  heated  by 
the  usual  American  convection  methods  (warm  air,  central  fan  and  direct 
hot  water  and  steam  heating  systems)  in  which  the  difference  between  the 
air  and  wall  surface  temperatures  may  not  be  great.  The  chart  does  not 
apply  to  rooms  heated  by  radiant  methods  such  as  the  British  panel 
system,  open  coal  fires,  and  the  like.  It  will  probably  not  apply  with 
adequate  accuracy  to  races  other  than  the  white  or  perhaps  to  inhabi- 
tants of  other  countries  where  the  living  conditions,  climate,  heating 
methods,  and  clothing  are  materially  different  from  those  of  the 
subjects  employed  in  experiments  at  the  Research  Laboratory. 

If  an  occupant  of  a  room  loses  heat  by  radiation  to  large  wall  or  glass 
surfaces  at  lower  temperatures,  the  air  within  the  room  must  be  main- 
tained at  a  higher  temperature  to  compensate  for  this  effect  in  order  to 
give  the  same  feeling  of  warmth.  The  results  of  a  recent  study 1&  by  the 
A.S.H.V.E.  Laboratory,  shown  in  Fig.  2,  indicate  that  in  po6rly  insulated 
buildings  this  effect  may  become  of  considerable  importance.  Thus  an 
occupant  of  a  room  having  inside  wall  surface  temperatures  of  55  F  on 
three  sides  will  require  an  air  temperature  of  7*4  F  to  have  the  same  feeling 
of  warmth  he  would  experience  in  at  warm-wall  room  with  air  at  70  F.  A 
wall  consisting  of  8-in.  brick  and  plaster,  with  16  F  outside  air  tenipera- 


*8Cold  Walls  and  Their  Relation  to  the  Feeling  of  Warmtfai  by  F.  C;  Hbtfefaffefc  an^Pau!  McDermott 
(A.S.H.V.E.  Journal  Section,  Heating,  Piping  and  Air  C&ndiiiomngt  JaBtfeftv'ldSS;  p:  53);    '  -      -  .      . 

41 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

ture  and  70  F  inside  air  temperature,  will  have  an  inside  surface  tem- 
perature of  55  F.  The  reverse  effect  will  be  experienced  by  occupants  of 
rooms  having  extensive  high-temperature  surfaces  in  them.  In  ^such 
cases,  a  lower  air  temperature  is  required  to  compensate  for  heat  radiated 
to  the  occupant. 

The  effective  temperature  index  for  persons  doing  medium  or  heavy 
muscular  work,  in  still  air,  has  also  been  determined  at  the  A.S.H.V.E. 
Research  Laboratory19. 


WALL   TEMPERATURE  DEC.  FAHR. 


FIG,  2.    CORRECTION  TO  VARIOUS  DRY-BULB  TEMPERATURES  IN  A  WARM- WALL  ROOM 
FOR  THE  SAME  FEELING  OF  WARMTH  IN  ROOMS  HAVING  THREE  COLD  WALLS. 
TEMPERATURES  INDICATED  BY  SHIELDED  THERMOMETERS  30  IN.  ABOVE  THE  FLOOR 


OPTIMUM  AIR  CONDITIONS 

No  single  comfort  standard  can  be  laid  down  which  would  meet  every 
need.  There  is  an  inherent  individual  variation  in  the  sensation  of 
warmth  or  comfort  felt  by  persons  when  exposed  to  an  identical  atmos- 
pheric condition.  The  state  of  health,  age,  sex,  clothing,  activity,  and 
the  degree  of  acquired  adaptation  seem  to  be  the  important  factors 
affecting  the  comfort  standards. 

Since  the  prolonged  effects  of  temperature,  humidity  and  air  move- 
ment on  health  are  not  known  to  the  same  extent  as  their  effects  on  com- 
fort, the  optimum  conditions  for  health  may  not  be  identical  with  those 
for  comfort.  On  general  physiologic  grounds,  however,  the  two  do  not 
differ  greatly  since  this  is  in  accordance  with  the  efficient  operation  of  the 
heat  regulating  mechanism  of  the  body.  This  belief  is  strengthened  by 


^Effective  Temperature  for  Persons  Lightly  Clothed  and  Working  in  Still  Air,  by  F.  C.  Houghten. 
W.  W.  Teague  and  W*  E.  Miller  (A.S.H.V.E.  TRANSACTIONS,  Vol.  32, 1926). 

42 


CHAPTER  2 — VENTILATION  AND  AIR  CONDITIONING  STANDARDS 

results  of  studies  on  premature  infants  over  a  four-year  period20.  By 
adjusting  the  temperature  and  humidity  so  as  to  stabilize  the  body  tem- 
perature of  these  infants,  the  incidence  of  diarrhoea  and  mortality  was 
decreased,  gains  in  body  weight  increased  and  infections  were  reduced 
to  a  minimum. 

Comfort  Chart;  Comfort  Line;  Comfort  Zone 

Fig.  3  shows  a  comfort  chart,  developed  at  the  A.S.H.V.E.  Laboratory, 
on  which  the  average  and  extreme  comfort  zones  have  been  superimposed. 
The  extreme  comfort  zone  includes  air  conditions  in  which  one  or  more  of 
the  experimental  subjects  were  comfortable.  The  average  comfort  zone 
includes  those  air  conditions  in  which  the  majority  of  the  subjects  (50  per 
cent  or  more)  were  comfortable.  That  particular  effective  temperature 
at  which  the  maximum  number  of  subjects  was  comfortable  was  called 
the  comfort  line. 

The  average  winter  comfort  zone  as  determined  at  the  A.S.H.V.E. 
Laboratory  ranges  from  63  deg  to  71  deg  ET  (effective  temperature). 
In  winter  while  at  rest,  a  large  percentage  of  persons  normally  clothed 
were  found  to  be  comfortable  at  66  deg  ET  and  this  temperature  has  been 
accepted  by  a  committee  of  the  Society21  as  the  winter  comfort  line  or 
optimum  effective  temperature. 

The  comfort  line  separates  the  cool  air  conditions  to  its  left  from  the 
warm  air  conditions  to  its  right.  Under  the  air  conditions  existing  along 
or  defined  by  the  comfort  line,  the  body  is  able  to  maintain  thermal 
equilibrium  with  its  environment  with  the  least  conscious  sensation  to  the 
individual,  or  with  the  minimum  phsyiologic  demand  on  the  heat  regulat- 
ing mechanism.  This  environment  involves  not  only  the  condition  of  the 
air  with  respect  to  temperature  and  humidity,  but  also  the  condition  of 
the  surrounding  objects  and  wall  surfaces.  The  comfort  zone  tests  were 
made  in  rooms  with  wall  surface  temperatures  approximately  the  same  as 
the  room  dry-bulb  temperature.  For  walls  of  large  area  having  unusually 
high  or  low  surface  temperatures,  however,  a  somewhat  lower  or  higher 
range  of  effective  temperature  is  required  to  compensate  for  the  increased 
gain  or  loss  of  heat  to  or  from  the  body  by  radiation22. 

The  average  summer  comfort  zone  for  exposures  of  3  hours  or  more 
ranges  from  about  66  deg  to  75  deg  ET,  based  on  studies  made  at  the 
Harvard  School  of  Public  Health17.  The  probable  optimum  effective 
temperature  (for  exposures  of  3  hours  or  more)  is  71  deg.  These  effective 
temperatures  average  about  4  deg  higher  than  those  found  in  winter  when 
customary  winter  clothing  was  worn.  The  variation  from  winter  to 
summer  is  probably  due  partly  to  adaptation  to  seasonal  weather  and 
partly  to  differences  in  the  clothing  worn  in  the  two  seasons. 

The  best  effective  temperature  (for  exposures  lasting  3  hours  or  more) 
was  found  to  follow  the  average  monthly  outdoor  temperature  more 
closely  than  the  prevailing  outdoor  temperature.  It  remained  at  approxi- 


»Applkation  of  Air  Conditioning  to  Premature  Nurseries  in  Hospitals,  by  C.  P.  Yagkra,  Philip  Drinker 
and  K.  D.  Blactfan  (A.S.H.V.E.  TRANSACTIONS,  VoL  36,  1930). 

»How  to  Use  the  Effective  Temperature  Index  and  Comfort  Charts  (A.S.H.V.E.  TRANSACTIONS, 
Vol.  38,  1932). 

»CoM  Walls  and  Their  Relation  to  the  Feeling  of  Warmth,  by  F.  C.  Houghten  and  Paul  McDermott 
(A.S.H.V.E.  Jomnal  Section,  Hea&ng*  Piping  and  Air  Conditioning,  January,  1933,  p.  53). 

43 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


90 


Air  Movement  or  Turbulence  15  to  25  ft.  per  mm. 


Average  Winter  Comfort  Zone  — - 
— ------   Average 'Winter  Comfort  Lme        | 

i','f,'f','\       Average  Summer'Comfort  Zone    | 
—  —  —  «••  Average  Summer  Comfort  Line 


70  80 

Pry  Bulb  Temperature  F 

FIG.  3.   A.S.H.V.E.  COMFORT  CHART  FOR  AIR  VELOCITIES  OF  15  TO  25  FPM  (STILL  AiR)2L 

Nate—  Both  summer  and  winter  comfort  zones  apply  to  inhabitants  of  the  United  States  only.  Applica- 
tion of  winter  comfort  line  is  further  limited  to  rooms  heated  by  central  station  systems  of  the  convection 
type.  The  line  does  not  apply  to  rooms  heated  by  radiant  methods.  Application  of  summer  comfort  line 
is  limited  to  homes,  offices  and  the  like,  where  the  occupants  become  fully  adapted  to  the  artificial  air  con- 
ditions- The  line  does  not  apply  to  theaters,  department  stores,  and  the  like  where  the  exposure  is  less  than 
3  hours. 


mately  the  same  value  in  July,  August  and  September,  and  although  the 
average  monthly  temperature  did  not  vary  much,  the  prevailing  outdoor 
temperature  ranged  from  70  F  to  99.5  F.  A  decrease  in  the  optimum 
temperature  became  apparent  only  when  the  prevailing  outdoor  tempera- 
ture fell  to  66  F,  which  is  below  the  customary  room  temperature  in  the 
United  States  for  summer  and  winter. 

,  Young  men  as  a  general  rule  prefer  conditions  in  the  cool  region  of  the 
comfort  zone,  and  women,  arid  older  people-, in  the  warm i region. of  the 

44 


CHAPTER  2 — VENTILATION  AND  AIR  CONDITIONING  STANDARDS 

comfort  zone.  Crowding  the  experimental  chamber  lowered  the  optimum 
effective  temperature  from  70.8  deg  when  the  gross  floor  area  per  occupant 
was  44  sq  ft  and  the  air  space  380  cu  ft,  to  69.4  deg  when  the  floor  area 
was  reduced  to  14  sq  ft  and  the  air  space  to  120  cu  ft  per  occupant. 

In  the  comfort  zone  experiments  of  the  A.S.H.V.E.  Research  Labora- 
tory, the  relative  humidity  was  varied  between  the  limits  of  30  and  70  per 
cent  approximately,  but  the  most  comfortable  range  has  not  been  deter- 
mined. In  similar  experiments  at  the  Harvard  School  of  Public  Health,  a 
relative  humidity  of  70  per  cent  was  found  to  be  somewhat  humid  in  winter, 
by  about  half  of  the  subjects  who  were  stripped  to  the  waist,  even  when 
the  dry-bulb  temperature  was  70  F  or  less.  In  summer,  a  relative  humi- 
dity of  30  per  cent  was  pronounced  as  a  little  too  dry  by  about  a  third  of 
the  subjects  wearing  warm-weather  clothing.  So  long  as  the  temperature 
was  kept  within  proper  limits,  the  majority  of  the  subjects  were  unable  to 
detect  sensations  of  humidity  (i.e.,  too  high,  too  low,  or  medium)  when 
the  relative  humidity  was  between  30  and  60  per  cent.  This  is  in  accord 
with  studies  by  Howell23,  Miura24  and  others. 

Dry  air  produces  an  excessive  loss  of  moisture  from  the  skin  and  respira- 
tory tract.  Owing  to  the  cooling  effect  of  evaporation,  higher  tempera- 
tures are  necessary,  and  this  condition  leads  to  discomfort  and  lassitude. 
Moist  air,  on  the  other  hand,  interferes  with  the  normal  evaporation  of 
moisture  from  the  skin,  and  again  may  cause  a  feeling  of  oppression  and 
lassitude,  especially  when  the  temperature  is  also  high. 

Just  what  the  optimum  range  of  humidity  is,  is  a  matter  of  conjecture. 
There  seems  to  exist  a  general  opinion,  supported  by  some  experimental 
and  statistical  data,  that  warm,  dry  air  is  less  pleasant  than  air  of  a 
moderate  humidity,  and  that  it  dries  up  the  mucous  membranes  in  such 
a  way  as  to  increase  susceptibility  to  colds  and  other  respiratory  dis- 
orders25- 26-  27. 

For  the  premature  infant,  a  high  relative  humidity  of  about  65  per  cent 
is  demonstrably  beneficial  to  health  and  growth28,  and  according  to 
Huntingdon29,  this  seems  to  be  the  case  for  adults  also.  All  of  these 
studies  indicate  that  the  optimum  humidity  must  always  be  considered 
in  combination  with  temperature. 

Until  more  exact  information  is  secured,  it  would  be  desirable  to  restrict 
the  comfort  zones  to  the  range  of  relative  humidity  employed  in  the 
comfort  zone  experiments,  namely,  30  to  70  per  cent.  Relative  humidities 
below  30  per  cent  may  prove  satisfactory  from  the  standpoint  of  comfort, 
so  long  as  extremely  low  humidities  are  avoided.  From  the  standpoint  of 
health,  however,  the  consensus  seems  to  favor  a  relative  humidity  between 

^Humidity  and  Comfort,  by  W.  H.  Howell  (The  Science  Press,  April,  1931). 

^Effect  of  Variation  in  Relative  Humidity  upon  Skin  Temperature  and  Sense  of  Comfort,  by  U.  Miura 
(American  Journal  of  Hygiene,  Vol.  13,  1931,  p.  432). 

^Reactions  of  the  Nasal  Cavity  and  Post-Nasal  Space  to  Chilling  of  the  Body  Surface,  by  Mudd,  Stuart, 
et  a!  (Journal  Experimental  Medicine,  1921,  Vol.  34,  p.  11). 

"Reactions  of  the  Nasal  Cavity  and  Post-Nasal  Space  to  Chilling  of  the  Body  Surfaces,  by  A.  Goldman, 
et  al  and  Concurrent  Study  of  Bacteriology  of  Nose  and  Throat  (Journal  Infectious  Diseases,  1921,  Vol.  29, 
p.  151). 

^The  Etiology  of  Acute  Inflammations  of  the  Noset  Pharynx  and  Tonsils,  by  Mudd,  Stuart,  et  al  (Am. 
Otol.,  RhinoL,  and  Laryngol.,  1921). 

^Application  of  Air  Conditioning  to  Premature  Nurseries  in  Hospitals,  by  C.  P.  Yaglou,  Philip  Drinker 
and  K.  D.  Blackfan  (A.S.H.V.E.  TRANSACTIONS,  VoL  36,  1930). 

^Weather  and  Health,  by  Ellsworth  Huntington  (Bulletin  of  the  National  Research  Council  No.  75. 
The  National  Academy  of  Science,  Washington,  D.  C.,  1930). 

45 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

40  and  60  per  cent.  In  mild  weather  such  comparatively  high  relative 
humidities  are  entirely  feasible,  but  in  cold  or  sub-freezing  weather  they 
are  objectionable  on  account  of  condensation  and  frosting  on  the^  windows. 
They  may  even  cause  serious  damage  to  certain  building  materials  of  the 
exposed  walls  by  condensation  and  freezing  of  the  moisture  accumulating 
inside  these  materials.  Unless  special  precautions  are  taken  to  properly 
insulate  the  affected  surfaces,  it  will  be  necessary  to  reduce  the  degree  of 
artificial  humidification  in  sub-freezing  weather  to  less  than  40  per  cent, 
according  to  the  outdoor  temperature.  Information  on  the  prevention  of 
condensation  on  building  surfaces  is  given  in  Chapter  7.  The  principles 
underlying  humidity  requirements  and  limitations  are  discussed  more 
fully  elsewhere30. 

The  comfort  chart  (Fig.  3)  applies  to  adults  between  20  and  70  years 
of  age  living  in  the  northeastern  parts  of  the  United  States.  For  pre- 
maturely born  infants,  the  optimum  temperature  varies  from  100  F  to 
75  F,  depending  upon  the  stage  of  development.  The  optimum  relative 
humidity  for  these  infants  is  placed  at  65  per  cent.  ^  No  data  are  yet 
available  on  the  optimum  air  conditions  for  full  term  infants  and  young 
children  up  to  school  age.  Satisfactory  air  conditions  for  these  age 
groups  are  assumed  to  vary  from  75  F  to  68  F  with  natural  indoor  humidi- 
ties. For  school  children,  the  studies  of  the  New  York  State  Commission 
on  Ventilation  place  the  optimum  air  conditions  at  66  F  to  68  F  tempera- 
ture with  a  moderate  humidity  (not  specified)  and  a  moderate  but  not 
excessive  amount  of  air  movement  (not  specified)31. 

Satisfactory  comfort  conditions  are  found  to  vary  from  40  deg  to  70  deg 
ET,  depending  upon  the  rate  of  work  and  amount  of  clothing  worn.  The 
effective  temperatures  giving  maximum  comfort  for  persons  working  have 
been  determined  by  the  A.S.H.V.E.  Research  Laboratory32  for  a  rate  of 
work  which  is  considered  hard  labor.  For  this  degree  of  work,  50  per  cent 
were  fairly  comfortable  for  temperatures  ranging  from  46  to  64  deg  ET, 
while  the  greatest  percentage  found  maximum  comfort  at  53  deg  ET. 
In  hot  industries,  80  deg  ET  is  considered  the  upper  limit  compatible 
with  efficiency,  and,  whenever  possible,  this  should  be  reduced  to  70  deg 
ET  or  less. 

APPLICATION  OF  COMFORT  CHART 

The  average  winter  comfort  line  (66  deg  ET)  applies  to  average 
American  men  and  women  living  inside  the  broad  geographic  belt  across 
the  United  States  in  which  central  heating  of  the  convection  type  is 
generally  used  during  four  to  eight  months  of  the  year.  It  does  not  apply 
to  rooms  heated  by  radiant  energy,  or  to  rooms  with  excessive  glass  area 
or  rooms  with  poorly  insulated  or  cold  walls,  and  it  has  not  been  advocated 
officially  for  use  in  foreign  countries  where  the  climate,  heating  methods, 
and  general  living  conditions  are  materially  different  from  those  in  the 
United  States,  although  several  foreign  workers  have  attempted  to  show 
that  it  cannot  be  so  applied.  Even  in  the  warm  south  and  southwestern 

*>Humidiiication  for  Residences,  by  A.  P.  Kratz  (University  of  Illinois  Engineering  Experiment  Station 
Bulletin  No.  230,  July  28,  1931). 

^Ventilation,  Report  of  the  New  York  State  Commission  on  Ventilation,  1923. 

»A.S.H.V.E.  research  paper  entitled  Heat  and  Moisture  Losses  from  Men  at  Work  and  Application  to 
Air  Conditioning'  Problems,  by  F,  C.  Houghten,  W.  W.  Teague,  W.  E.  Miller  and  W.  P.  Yant  (A.S.H'.V.E. 
TRANSACTIONS,  Vol.  37,  1931), 

46 


CHAPTER  2 — VENTILATION  AND  AIR  CONDITIONING  STANDARDS 

climates,  and  in  the  very  cold  north-central  climate  of  the  United  States, 
the  comfort  chart  would  probably  have  to  be  modified  according  to 
climate,  living  and  working  conditions,  and  the  degree  of  acquired 
adaptation. 

In  densely  occupied  spaces,  such  as  classrooms,  theaters  and  audi- 
toriums, somewhat  lower  temperatures  are  necessary  than  those  indicated 
by  the  comfort  line  on  account  of  counter-radiation  between  the  bodies  of 
occupants22  in  close  proximity.  In  rooms  in  which  the  average  wall 
surface  temperature  is  considerably  below  the  air  temperature,  higher  air 
temperatures  are  necessary.  The  reverse  holds  true  in  radiant  or  panel 
heating  methods.  (See  Chapter  38.) 

The  sensation  of  comfort,  in  so  far  as  the  physical  environment  is  con- 
cerned, is  not  absolute  but  varies  considerably  among  certain  individuals. 
Therefore,  in  applying  the  air  conditions  indicated  by  the  comfort  line, 
it  should  not  be  expected  that  all  the  occupants  of  a  room  will  feel  per- 
fectly comfortable.  When  the  winter  comfort  line  is  applied  in  accordance 
with  the  foregoing  recommendations,  the  majority  of  the  occupants  will 
be  perfectly  comfortable,  but  there  will  always  be  a  few  who  would  feel 
a  bit  too  cool  and  a  few  a  bit  too  warm.  These  individual  differences  among 
the  minority  should  be  counteracted  by  suitable  clothing. 

Air  conditions  lying  outside  the  average  comfort  zone  but  within  the 
extreme  comfort  zone  may  be  comfortable  to  certain  persons.  In  other 
words,  it  is  possible  for  half  of  the  occupants  of  a  room  to  be  comfortable 
in  air  conditions  outside  the  average  comfort  zone,  but  in  the  majority  of 
cases,  if  not  in  all,  these  conditions  will  be  well  within  the  extreme  comfort 
zone  as  determined  experimentally. 

Strictly  speaking,  the  only  authoritative  comfort  zone  on  which  accur- 
ate data  are  available,  is  that  for  15  to  25  fpm  air  movement  or  tur- 
bulance  (often  referred  to  as  still  air).  In  the  past,  the  winter  comfort 
zone  has  often  been  superimposed  on  the  thermometric  chart  or  on  effec- 
tive temperature  charts  for  various  air  velocities,  on  the  assumption  that 
air  conditions  of  equal  warmth  are  approximately  equally  comfortable. 
This  may  hold  in  hot  industries  where  the  workers  are  adapted  to  high 
temperatures  and  strong  air  currents,  but  it  does  not  apply  to  sedentary 
conditions.  To  ascertain  approximately  whether  a  given  industrial  con- 
dition is  reasonably  comfortable,  it  would  be  necessary  first  to  compute 
the  effective  temperature  from  the  thermometric  chart  (Fig.  1)  and  then 
to  refer  this  effective  temperature  to  the  comfort  chart  (Fig.  3),  or  to 
refer  directly  to  a  chart  or  table  for  the  proper  air  velocity. 

The  summer  comfort  line  (71  deg  ET)  is  applicable  to  the  same  geo- 
graphic area  as  the  winter  comfort  line.  It  is  further  restricted  to  cases  in 
which  the  human  body  has  reached  thermal  equilibrium  with  its  environ- 
ment. As  a  general  rule  this  takes  place  after  1}^  to  3  hours'  exposure. 
When  a  person  from  outdoors  enters  a  room  cooled  to  71  deg  ET  on  a  hot 
day  (95  F  or  over)  an  intense  chill  is  likely  to  be  experienced  which  is 
unpleasant.  However,  after  remaining  in  the  room  for  about  2  hours, 
this  fundamental  optimum  condition  will  prove  satisfactory  to  the  average 
person.  The  summer  comfort  zone,  as  well  as  the  comfort  line,  makes 
proper  allowance  for  these  adaptive  changes  in  the  body,  and  thus  applies 
to  homes,  offices,  schools  and  other  similar  places  where  persons  of 
sedentary  occupations  speed  from  3  to  8  or  more  hours  daily. 

47 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

In  artificially  cooled  theaters,  department  stores,  restaurants,  and  other 
public  buildings  where  the  period  of  occupancy  is  short,  the  contrast 
between  outdoor  and  indoor  air  conditions  becomes  the  deciding  factor  in 
regard  to  the  temperature  and  humidity  to  be  maintained.  The  object  of 
cooling  such  places  in  the  summer  is  not  to  reduce  the  temperature  to  the 
optimum  degree,  but  to  maintain  therein  a  temperature  which  is  tem- 
porarily comfortable  to  the  patrons  who  thus  avoid  sensations  of  chill  and 
intense  heat  on  entering  and  leaving  the  building.  The  relative  humidity . 
should  be  low  enough  (about  50  per  cent  or  less)  to  give  a  sense  of  comfort 
without  chill  and  to  induce  a  rate  of  evaporation  which  will  keep  clothing 
and  skin  dry.  For  exposures  less  than  3  hours,  desirable  indoor  conditions 
in  summer  corresponding  to  various  outdoor  temperatures  are  given  in 
Table  2. 

It  should  be  kept  in  mind  that  southern  people,  with  their  more  sluggish 
heat  production  and  lack  of  adaptability,  will  demand  a  comfort  zone 
several  degrees  higher  than  those  given  here  for  the  more  active  people  of 


TABLE  2.    DESIRABLE  INDOOR  AIR  CONDITIONS  IN  SUMMER  CORRESPONDING 
TO  OUTDOOR  TEMPERATURES 

Applicable  to  Exposures  Less  Than  3  Hours 


OUTDOOR  TEMPERATURE 
(!>EG  FAHB) 

INDOOR  Am  CONDITIONS  WITH  DEW  POINT 
CONSTANT  AT  57  F 

DHT-BTJLB 

DET-BTJLB 

WET-BULB 

EFFECTIVE  TEMP 

95 

80.0 

65.0 

73 

90 

78.0 

64.5 

72 

85 

76.5 

64.0 

71 

80 

75.0 

63.5 

70 

75 

73.5 

63.0 

69 

70 

72.0 

62.5 

68 

northern  climates.  Instead  of  the  summer  comfort  line  standing  at 
71  deg  as  here  given,  it  was  found  to  be  much  higher  for  foreigners  in 
Shanghai  where  climatic  conditions  are  similar  to  those  of  our  gulf 
states.  This  difference  in  basic  metabolic  level  of  people  forms  a  very 
real  problem  for  air  conditioning  engineers,  which  they  must  recognize  in 
their  efforts  to  give  proper  conditions  of  comfort.  Cooling  of  theaters, 
resturants,  and  other  public  buildings  in  southern  climates  cannot  be 
based  on  northern  standards  without  considerable  modification. 


A.S.H.V.E.  VENTILATION  STANDARDS33 

It  is  the  intent  of  the  Committee  in  presenting  this  report  to  confine  itself  to  a  statement  of  those 
requirements  which,  based  on  present  day  knowledge,  will  provide  adequate  ventilation  for 
spaces  intended  for  human  occupancy.  The  following  standards  shall  apply  to  all  spaces 
occupied  by  human  beings  in  all  buildings  for  which  ventilation  regulations  are  to  be  established. 


^Report  of  A.S.H.V.E.  Committee  on  Ventilation  Standards  consisting  of  W.  H.  Driscoll,  Chairman, 
J.  J.  Aeberly,  F.  Paul  Anderson,  L.  A.  Harding,  D.  D.  Kimball,  J.  R.  McCoIl,  C.  L.  Riley,  W.  A.  Rowe, 
Perry  West  and  A.  C.  Willard,  presented  at  the  Serai-Annual  Meeting  of  the  Society,  Milwaukee,  Wis., 
June,  1932,  and  adopted  by  the  Society  in  August,  1932. 

48 


CHAPTER  2 — VENTILATION  AND  AIR  CONDITIONING  STANDARDS 

SECTION   I— AIR  TEMPERATURE  AND  HUMIDITY 

The  temperature  and  humidity  of  the  air  in  such  occupied  spaces,  and  in  which  the  only 
source  of  contamination  is  the  occupant,  shall  be  maintained  at  all  times  during  occu- 
pancy at  an  Effective  Temperature,  as  hereinafter  stated. 

The  relative  humidity  shall  be  not  less  than  30  per  cent,  nor  more  than  60  per  cent  in 
any  case.  The  Effective  Temperature  shall  range  between  64  deg  and  69  deg  when 
heating  or  humidification  is  required,  and  between  69  deg  and  73  deg  when  cooling  or 
dehumidification  is  required. 

These  Effective  Temperatures  shall  be  maintained  at  a  level  of  36  in.  above  the  floor. 
(See  Appendix,  Tables  A  and  B). 

SECTION   II— AIR  QUALITY 

The  air  in  such  occupied  spaces  shall  at  all  times  be  free  from  toxic,  unhealthful  or 
disagreeable  gases  and  fumes  and  shall  be  relatively  free  from  odors  and  dust. 

In  every  space  coming  within  the  provisions  of  these  requirements  and  in  which  the 
quality  of  the  air  is  below  the  standards  prescribed  by  good  medical  and  engineering 
practices,  due  to  toxic  substances,  bacteria,  dust,  excessive  temperature,  excessive 
humidity,  objectionable  odors,  or  other  similar  causes,  means  for  ventilating  shall  be 
provided  so  that  the  quality  of  the  air  shall  be  raised  to  these  standards. 

SECTION  III— AIR  MOTION 

The  air  in  such  occupied  spaces  shall  at  all  times  be  in  constant  motion  sufficient  to 
maintain  a  reasonable  uniformity  of  temperature  and  humidity,  but  not  such  as  to  cause 
objectionable  drafts  in  any  occupied  portion  of  such  spaces. 

The  air  motion  in  such  occupied  spaces,  and  in  which  the  only  source  of  contamination 
is  the  occupant,  shall  have  a  velocity  of  not  more  than  50  feet  per  minute,  measured 
at  a  height  of  36  in.  above  the  floor. 

SECTION  IV— AIR  DISTRIBUTION 

The  air  in  all  rooms  and  enclosed  spaces  shall,  under  the  provisions  of  these  reqr're- 
ments,  be  distributed  with  reasonable  uniformity,  and  the  variation  in  the  carbon  dioxide 
content  of  the  air  shall  be  taken  as  a  measure  of  such  distribution. 

The  air  in  a  space  ventilated  in  accordance  with  these  requirements,  and  in  which  the 
only  source  of  contamination  is  the  occupant,  shall  be  distributed  and  circulated  so  that 
the  variation  in  the  concentration  of  carbon  dioxide,  when  measured  at  a  height  of 
36  in.  above  the  floor,  shall  not  exceed  one  part  in  10,000. 

SECTION  V— AIR  QUANTITY 

The  quantity  of  air  used  to  ventilate  the  given  space  during  occupancy  shall  always 
be  sufficient  to  maintain  the  standards  of  air  temperature,  air  quality,  air  motion  and  air 
distribution  as  herein  required.  Not  less  than  10  cubic  feet  per  minute  per  occupant  of 
the  total  air  circulated  to  meet  these  requirements  shall  be  taken  from  an  outdoor  source. 

APPENDIX 
Definitions 

For  the  purposes  of  these  standards  the  terms  used  shall  be  defined  as  follows: — 

Ventilation :  The  process  of  supplying  or  removing  air  by  natural  or  mechanical  means,  to  or  from  any 
space.  Such  air  may  or  may  not  have  been  conditioned.  (See  Air  Conditioning). 

Air  Conditioning:  The  simultaneous  control  of  all  or  at  least  the  first  three  of  those  factors  affecting 
both  the  physical  and  chemical  conditions  of  the  atmosphere  within  any  structure.  These  factors  include 
temperature,  humidity,  motion,  distribution,  dust,  bacteria,  odors,  toxic  gases,  and  ionization,  most  of 
which  affect  in  greater  or  lesser  degree  human  health  or  comfort. 

Dry-Bulb  Temperature:  The  temperature  of  the  air  which  is  indicated  by  any  type  of  thermometer 
which  is  not  affected  by  the  water  vapor  content  or  relative  humidity  of  the  air. 

Dust:  Solid  material  in  a  finely  divided  state,  the  particles  of  which  are  large  and  heavy  enough  to  fall 
with  increasing  velocity,  due  to  gravity  in  still  air.  For  instance,  particles  of  fine  sand  or  grit,  such  as  are 
blown  on  a  windy  day,  the  average  diameter  of  which  is  approximately  0.01  centimeter,  may  be  called  dust. 

Effective  Temperature:  An  arbitrary  index  of  the  degree  of  warmth  or  cold  felt  by  the  human  body 
in  response  to  temperature,  humidity,  and  movement  of  the  air.  Effective  temperature  is  a  composite 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

Index  which  combines  the  readings  of  temperature,  humidity,  and  air  motion  into  a  single  value.  The 
numerical  value  of  the  effective  temperature  scale  has  been  fixed  by  the  temperature  of  saturated  air  which 
induces  an  identical  sensation  of  warmth. 

Humidity:  The  water  vapor  (either  saturated  or  superheated  steam)  occupying  any  space,  which  may 
or  may  not  contain  other  vapors  and  gases  at  the  same  time. 

Relative  Humidity:  A  ratio,  although  usually  expressed  in  per  cent,  used  to  indicate  the  degree  of 
saturation  existing  in  any  given  space  resulting  from  the  water  vapor  present  in  that  space.  The  presence 
of  air  or  other  gases  in  the  same  space  at  the  same  time  has  nothing  to  do  with  the  relative  humidity  of 
the  space,  which  depends  merely  on  the  temperature  and  partial  pressure  of  the  vapor. 

Spaces  in  Which  the  Only  Source  of  Contamination  Is  the  Occupant:  Spaces  in  which  the 
atmospheric  contamination  results  entirely  from  the  respiratory  processes  of  the  occupant,  including  heat, 
moisture,  and  odors  given  off  by  the  body.  No  manufacturing  or  industrial  processes  or  other  sources  of 
atmospheric  contamination,  including  heat  and  moisture,  than  people  are  considered  under  this  title. 

TABLE  A.    EFFECTIVE  TEMPERATURES  RANGING  FROM  64  DEC  TO  69  DEC  FOR  VARIOUS  DRY-  BULB  TEM- 
PERATURES AND  RELATIVE  HUMIDITIES  FOR  STILL  AIR  FOR  PERSONS 
NORMALLY    CLOTHED    AND    SLIGHTLY 


(For  use  when  heating  or  humidification  is  required) 


RELATIVE  HUMIDITIES  (PER  CENT) 

DRY-  BULB 

TEMPERATURES 
(DEC  FAHR) 

30         ^         35 

40                  45 

50 

55 

60 

EFFECTIVE    TEMPERATURES    (DEGREES) 

67 

64.0 

64.3 

68 

64.0 

64.2 

64.5 

64.8 

65.1 

69 

64.1 

64.4 

64.8 

65.1 

65.4 

65.7 

66.0 

70 

64.8 

65.1 

65.4 

65.8 

66.2 

66.5 

66.8 

71 

65.5 

65.8 

66.2 

66.6 

67.0 

67.3 

67.7 

72 

66.2 

66.5 

66.9 

67.3 

67.7 

68.1 

68.5 

73 

67.0 

67.3 

67.7 

68.1 

68.5 

68.9 

74 

67.7 

68.0 

68.4 

68.8 

75 

68.4 

68.7 

76 

69.0 

aSee  Fig.  3. 


TABLE  B.    EFFECTIVE  TEMPERATURES  RANGING  FROM  69  DEC  TO  73  DEC  FOR  VARIOUS  DRY-BULB  TEM- 
PERATURES AND  RELATIVE  HUMIDITIES  FOR  STILL  AIR  FOR  PERSONS 
NORMALLY    CLOTHED    AND    SLIGHTLY    AcTivEa-b 

(For  use  when  cooling  or  dehumidification  is  required) 


RELATIVE  HUMIDITIES  (PER  CENT) 


DRY-BULB 

TEMPERATURES 
(DEG  FAHR) 

30 

35 

40 

45 

50 

55 

60 

EFFECTIVE    TEMPERATURES    (DEGREES) 

73 

69.3 

74 

69.3 

69.7 

70.1 

75 

69.1 

69.5 

70.0 

71.5 

71.0 

76 

69.0 

69.4 

69.9 

70.5 

70.8 

71.3 

71.8 

77 

69.7 

70.2 

70.7 

71.2 

71.6 

72.1 

72.6 

78 

70.4 

70.9 

71.4 

71.9 

72.4 

73.0 

79 

71.1 

71.6 

72.2 

72.6 

80 

71.8 

72.4 

72.9 

81 

72.5 

•See  Fig,  3. 

bThis  table  applies  primarily  to  cases  in  which  the  human  body  has  reached  equilibrium  with  the  sur- 
rounding air.  A  higher  plane  of  summer  effective  temperatures  is  required  in  places  of  public  assembly 
where  the  period  of  occupancy  is  short,  than  is  required  for  offices  and  industrial  plants  where  the  period  of 
occupaacy  isjrf  longer  duration.  When  the  period  of  occupancy  is  two  hours  or  less,  the  dry-bulb  tempera- 
ture shall  be  72  F  plus  one-third  of  the  difference  between  the  outside  dry-bulb  temperature  and  70  F,  and 
the  relative  humidity  shall  not  exceed  60  per  cent.  (See  also  Table  2.) 


50 


CHAPTER  2 — VENTILATION  AND  AIR  CONDITIONING  STANDARDS 

FACTORS  INFLUENCING  APPLICATIONS 

The  conditions  and  limitations  outlined  under  the  heading  Application 
of  Comfort  Chart  should  be  noted  in  applying  the  temperatures  and 
relative  humidities  specified  in  Tables  A  and  B  of  the  preceding 
A.S.H.V.E.  Ventilation  Standards. 

Air  Quality 

In  occupied  spaces  in  which  the  vitiation  is  entirely  of  human  origin, 
the  chemical  composition  of  the  air,  the  dust,  and  bacteria  content  may  be 
dismissed  from  consideration  so  that  the  problem  consists  in  maintaining 
a  suitable  temperature  with  a  moderate  humidity,  and  in  keeping  the 
atmosphere  free  from  objectionable  odors.  Such  unpleasant  odors, 
human  or  otherwise,  can  be  easily  detected  by  persons  entering  the  room 
from  clean,  odorless  air.  A  further  discussion  of  air  quality  will  be  found 
in  Chapters  15  and  16. 

Air  Motion 

As  a  result  of  studies  by  Baetjer34  and  work  carried  on  by  the  A.S.H.V.E. 
Research  Laboratory,  it  is  now  recognized  that  the  importance  of  air 
motion  in  air  conditioning  ranks  only  second  to  temperature.  Air  in  an 
occupied  space  having  all  the  other  essential  qualities  but  lacking  in  air 
motion  feels  stagnant,  stuffy,  and  depressing,  because  the  vitiated  air 
next  to  the  body  is  not  replaced  by  the  surrounding  air  possessing  the 
satisfactory  qualities.  Hence,  air  motion  is  absolutely  essential  that  an 
occupant  may  realize  the  other  desired  qualities  of  the  atmosphere. 
Possible  limits  in  variation  in  air  motion  may  range  from  5  fpm  to  50  fpm, 
as  measured  by  the  Kata  thermometer.  (See  Chapter  40.)  However, 
satisfactory  results  are  more  likely  to  be  insured  by  air  velocities  ranging 
from  15  to  30  fpm.  The  limit  of  5  fpm  may  be  taken  as  the  minimum 
during  the  heating  season,  and  50  fpm  as  the  maximum  for  the  cooling 
season. 

Air  Distribution 

Variation  in  concentration  of  carbon  dioxide  in  different  parts  of  an 
occupied  room  has  been  used  as  a  measure  of  satisfactory  distribution  of 
the  outside  or  conditioned  air  supply.  For  satisfactory  air  distribution, 
the  carbon  dioxide  concentration  at  the  36-in.  level  should  not  vary  by 
more  than  one  part  in  10,000  parts  of  air.  Recent  work2  by  the  A.S.H.V.E. 
Research  Laboratory  demonstrates  that  variations  in  dry-bulb  tempera- 
ture, wet-bulb  temperature,  or  moisture  content  of  the  air  are  equally 
good  indices  of  air  distribution.  This  work  also  indicates  that  the 
presence  of  satisfactory  air  motion  within  the  room  (15  to  30  fpm  as 
measured  by  the  Kata  thermometer)  insures  satisfactory  distribution. 
Because  of  the  laborious  and  exacting  technique  involved  in  making 
carbon  dioxide  determinations,  it  is  recommended  that  satisfactory 
distribution  can  be  amply  insured  by  the  presence  of  such  air  velocities  in 
all  parts  of  the  room  together  with  dry-bulb  temperature  variations  of  not 
to  exceed  3  deg  at  the  36-in.  level. 


^Threshold  Air  Currents  In  Ventilation  (American  Journal  of  Hygiene,  Vol.  IVr  No.  8,  p.  650,  1924). 

51 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

Air  Quantity 

The  quantity  of  air  to  be  circulated  through  an  occupied  space,  whether 
by  natural  or  mechanical  means,  or  whether  the  air  is  conditioned  or  not, 
must  in  all  cases  be  sufficient  to  maintain  the  required  standards  of  air 
temperature,  quality,  motion  and  distribution.  The  factors  which  deter- 
mine air  quantity  include  the  type  and  nature  of  the  building,  locality, 
climate,  height  of  rooms,  floor  area,  window  area,  extent  of  occupancy, 
and  last  but  not  least,  the  method  of  distribution. 

The  quantity  of  air  supplied  to  a  room  by  an  air  conditioning  or  venti- 
lating system  serves  two  purposes:  First,  the  supply  of  sufficient  outside 
air  for  the  needs  of  the  occupants;  and  second,  the  setting  up  of  circulation 
or  air  motion  within  the  room.  Until  recently  it  was  considered  that 
30  cfm  were  necessary  in  any  occupied  space,  particularly  in  a  classroom, 
It  has  since  been  demonstrated  that  10  cfm  of  outside  air  per  person  is 
frequently  sufficient  to  remove  body  heat,  insure  against  body  odors, 
and  provide  the  chemical  needs  of  respiration.  However,  it  is  found 
that  a  greater  volume  should  be  circulated  in  the  average  room  in 
order  to  provide  the  required  air  motion.  It  is  now  customary  to  supply 
the  minimum  amount  of  outside  or  conditioned  air  required  for  removing 
heat  and  odors,  and  to  recirculate  the  additional  volume. 

In  offices  and  small  rooms  where  the  occupants  smoke,  from  6  to  7  cfm 
of  outside  air  per  occupant  will  be  necessary  to  eliminate  the  nuisance 
effects  of  the  smoke ;  this  quantity  of  air,  however,  may  be  a  part  of  that 
necessary. for  other  ventilation  requirements.  Restaurants  which  permit 
smoking,  because  of  the  exposed  food  and  the  necessity  that  restaurant 
air  seem  very  clean,  need  from  10  to  12  cfm  of  outside  air  per  occupant  to 
care  for  the  smoke  condition.  This  air,  likewise,  need  not  be  in  addition 
to  that  required  for  other  ventilation  purposes. 

Temperature  Rise 

The  total  quantity  of  air  introduced  is  governed  largely  by  the  needs 
for  controlling  temperature  and  humidity  when  either  heating  or  cooling 
is  required.  As  a  rule,  the  introduction  and  distribution  of  warm  air  into 
an  occupied  space  does  not  present  as  many  difficulties  as  does  the  intro- 
duction of  cold  air.  The  former  is  determined  from  the  amount  of  heat  to 
be  given  up  to  the  space,  and  the  latter  is  determined  from  the  amount  of 
heat  to  be  removed  from  the  space,  using  a  temperature  rise  that  will 
produce  uniform  distribution  without  the  production  of  disagreeable 
drafts. 

Fig.  4  shows  the  changes  in  carbon  dioxide  concentration  and  moisture 
content  resulting  from  occupation,  in  the  atmosphere  of  a  room  supplied 
with  various  volumes  of  outside  air.  Data  are  given  for  an  adult,  5  ft 
8  in.  in  height  weighing  150  pounds  and  having  a  body  surface  area  of 

19.5  sq  ft,  and  for  a  child,  12  years  of  age,  4  ft  7  in.  in  height,  weighing 

76.6  pounds  and  having  a  body  surface  area  of  12.6  sq  ft.    It  is  a  recognized 
fact  that  the  dissipation  of  heat  and  moisture  to  the  atmosphere,  the 
addition  of  carbon  dioxide,  and  all  metabolic  changes  take  place  in  pro- 
portion to  the  surface  area  of  the  individual.    Hence,  data  for  persons  of 
other  sizes  may  be  obtained  by  interpolating  among  the  curves  given. 
The  rate  of  sensible  heat  production  is  given  in  Fig.  7.    Fig.  4  also  gives 

52 


CHAPTER  2 — VENTILATION  AND  AIR  CONDITIONING  STANDARDS 

the  temperature  of  the  incoming  air  necessary  to  maintain  a  room  tem- 
perature of  either  70  or  80  F  as  indicated,  assuming  that  there  is  no  heat 
gain  or  loss  to  the  room  by  transmission  through  the  walls,  solar  radiation 
or  other  sources. 


ADULTS     IN     63  F     TO     86  F     AIR 


CHILDREN     IN     63  F     TO 


ADULTS     IN     8OF    AIR 


CHILDREN     IN    8O  F    AIR 


ADULTS     IN     70  F     AIR 


CHILDREN     IN     70  F    AIR 


CHILDREN     IN    80  F    AIR- 


ULTS    IN    80 F    AIR 


CHILDREN     IN     70  F    AIR 


ADULTS     IN     70  F     AIR 


12  16  20 

RATE     OF     AIR    SUPPLY 
CUBIC    FEET    PER    MINUTE    PER     OCCUPANT 


24 


FIG.  4.    RELATION  AMONG  RATE  OF  AIR  CHANGE  PER  OCCUPANT,  CARBON  DIOXIDE 

CONCENTRATION  AND  MOISTURE  CONTENT  OF  ENCLOSURE,  AND  DRY-BULB 

TEMPERATURE  OF^  INCOMING  AIR 

Two  of  the  most  important  factors  on  which  the  temperature  rise 
depends  are  (1)  the  method  of  distribution  and  (2)  the  most  economical 
temperature  rise  for  the  conditions  involved.  Some  systems  of  distri- 

53 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

bution  produce  drafts  with  but  a  few  degrees  temperature  rise,  while 
other  systems  operate  successfully  with  a  temperature  rise  as  high  as 
35  deg.  The  total  air  quantity  introduced  in  any  particular  case  is 
inversely  proportional  to  the  temperature  rise,  and  depends  largely  upon 
the  judgment  and  ingenuity  of  the  engineer  in  designing  the  most  suitable 
system  for  the  particular  conditions.  Small  quantities  of  air  reduce  the 
size  of  equipment,  ducts,  space,  and  initial  cost,  but  require  lower  air 
temperatures.  In  any  specific  case,  the  cost  of  refrigeration  must  be 
balanced  against  the  extra  cost  in  increased  size  of  equipment  and 
running  expense. 

Outside  Air.  In  order  to  provide  uniform  temperature  conditions,  it 
is  necessary  to  maintain  a  pressure  of  about  0.1  in.  of  water  in  the  room  or 
space  to  be  ventilated  or  conditioned.  This  usually  requires  the  intro- 
duction of  a  certain  amount  of  outside  air  which  depends  on  the  particular 
conditions  involved,  and  may  vary  over  a  considerable  range. 

In  rooms  in  which  the  only  source  of  contamination  is  the  occupant  the 
minimum  quantity  of  outside  or  new  air  to  be  circulated  appears  to  be 
that  necessary  to  remove  objectionable  body  odors.  The  concentration  of 
body  odors  in  turn  depends  largely  upon  the  temperature  of  the  air ;  the 
higher  the  temperature,  the  greater  the  amount  of  perspiration  (sensible 
or  insensible)  given  off  from  the  skin,  and  the  greater  the  concentration 
of  odors. 

NATURAL  AND  MECHANICAL  VENTILATION 

Under  favorable  conditions  natural  ventilation  methods  properly 
combined  with  means  for  heating  may  be  sufficient  to  provide  for  the 
foregoing  standards.  As  a  rule,  in  instances  in  which  the  only  source  of 
contamination  is  the  occupant,  the  requirements  may  be  fulfilled  when 
the  following  conditions  prevail: 

1.  At  least  50  sq  ft  of  floor  area  for  each  occupant. 

2.  At  least  500  cu  ft  of  air  space  per  occupant. 

3.  Effective  openings  in  windows  and  skylights  equal  to  at  least  5  per  cent  of  the 
floor  area. 

Whenever  natural  means  are  not  sufficient  to  maintain  the  standards, 
resort  must  be  made  to  whatever  modifications  or  mechanical  apparatus 
are  necessary  to  secure  such  standards. 

In  large  offices,  large  school  rooms,  and  in  public  and  industrial  build- 
ings, natural  ventilation  is  uncertain  and  makes  heating  difficult.  The 
chief  disadvantage  of  natural  methods  is  the  lack  of  control :  they  depend 
largely  on  weather  and  upon  the  velocity  and  direction  of  the  wind. 
Rooms  on  the  windward  side  of  a  building  may  be  difficult  to  heat  and 
ventilate  on  account  of  drafts,  while  rooms  on  the  leeward  side  may  not 
receive  an  adequate  amount  of  air  from  out  of  doors.  The  partial  vacuum 
produced  on  the  leeward  side  under  the  action  of  the  wind  may  even 
reverse  the  flow  of  air  so  that  the  leeward  half  of  the  building  has  to  take 
the  drift  of  the  air  from  the  rooms  of  the  windward  half.  Under  such 
conditions  no  outdoor  air  would  enter  through  a  leeward  window  opening, 
but  room  air  would  pass  out. 

In  warm  weather  natural  methods  of  ventilation  afford  little  or  no 

54 


CHAPTER  2 — VENTILATION  AND  AIR  CONDITIONING  STANDARDS 


control  of  indoor  temperature  and  humidity.    Outdoor  smoke,  dust  and 
noise  constitute  other  limitations  of  natural  methods. 

REC1RCULATION 

The  saving  in  operating  costs  due  to  recirculation  of  the  air,  while  very 
considerable,  must  not  be  obtained  at  the  expense  of  air  quality.  The 
percentage  of  recirculated  air  may  be  varied  to  suit  the  seasonal  changes 
so  as  to  conserve  heat  in  winter  and  refrigeration  in  summer,  but  at  no 
time  during  occupancy  should  there  be  taken  from  out  of  doors  less  than 
10  cfm  for  each  ^  occupant.  ^As  a  general  rule,  recirculation  impairs  the 
quality  of  the  air  by  excessive  humidity  (if  not  conditioned),  excessive 
odors,  or  both,  and  it  tends  to  deprive  the  air  of  its  ionic  content,  but 


77/ff 

FIG.  5.     INFLUENCE  OF  ROOM  OCCUPANCY  ON  IONIC  CONTENT^ 

(Cubical  Contents  of  Room,  10,000  Cu  Ft;  Number  of  Occupants,  34) 

the  influence  of  this  factor  on  comfort  and  health  is  at  present  a  matter 
of  speculation. 

Toilets,  kitchens,  and  similar  rooms,  in  buildings  using  recirculation, 
should  be  separately  mechanically  ventilated  by  exhausting  the  air  from 
them  in  order  to  prevent  objectipnable  odors  from  diffusing  into  other 
parts  of  the  building. 

ULTRA-VIOLET  RADIATION  AND  IONIZATION 

In  spite  of  the  rapid  advances  made  in  the  field  of  air  conditioning 
during  the  past  few  years,  the  secret  of  reproducing,  in  indoor  spaces, 
atmospheres  of  as  stimulating  qualities  as  those  existing  outdoors  in  the 
country,  under  ideal  weather  conditions,  has  not  as  yet  been  found.  In 
fact,  extensive  studies  have  failed  to  elucidate  the  cause  of  the  stimulating 
quality  of  outdoor  country  air,  qualities  which  are  lost  when  such  air  is 
brought  indoors  and  particularly  when  it  is  handled  by  mechanical 

55 


AMERICAS  SOCIETY  of  HEATIKG  and  VENTILATING  ENGINEERS  GUIDE,  1935 
TABLE  3.    RELATION  BETWEEN  METABOLIC  RATE  AND  ACTIVITY* 


A.C7IVITT 

METABOLIC  RATE  Brtr 
PEB  HOTTR  FOR  AVERAGE 
MAN  (19-5  SQ  FT  SUB- 
FACE  AREA) 

ACTHORITT 

Seated  at  rest 

384 

Research  Laboratory,  American  Society  of 

Standing  at  rest 

431 

Heating  and  Ventilating  Engineers. 
Research  Laboratory,  American  Society  of 

Walking  2  mph  

761 

Heating  and  Ventilating  Engineers. 
Average    values    from     Douglas,    Haidane, 

Walking  3  rnph 

1049 

Henderson  and  Schneider;  and   Henderson 
and  Haggard. 
Douglas,  Haldane,  Henderson  and  Schneider 

Walking  4  mph  

1388 

Average   values    from     Douglas,    Haldane, 

\Valking  5  mph 

2530 

Henderson  and  Schneider;  and  Henderson 
and  Haggard. 
Douglas,  Haldane,  Henderson  and  Schneider 

Slow  run     .        

2285 

Henderson  and  Haggard 

Very  severe  exercise-..  . 
Maximum  exertion  
Tailor    

2555 
3333  to  4762  -h 
482 

Benedict  and  Carpenter 
Henderson  and  Haggard 
Becker  and  Hamalainen 

Bookbinder 

626 

Becker  and  Hamalainen 

Shoemaker            .    .. 

661 

Becker  and  Hamalainen 

Carpenter 

762  to  963 

Becker  and  Hamalainen 

Metal  worker    

862 

Becker  and  Hamalainen 

Painter  (of  furniture).. 
Stonemason  

876 
1488 

Becker  and  Hamalainen 
Becker  and  Hamalainen 

Man  sawing  wood 

1797 

Becker  and  Hamalainen 

means.  It  is  true  that  many  suggestions  have  been  advanced  to  account 
for  the  stimulating  quality  of  outdoor  air,  such  as  ultra.- violet  light  _and 
ionization.  At  the  present  time  neither  of  these  suggestions  has  received 
any  degree  of  scientific  confirmation. 

It  is  generally  recognized  that  total  outdoor  solar  radiation  has  marked 
curative  value  in  certain  diseases  and  is  also  a  powerful  germicidal  agent. 
A  critical  review  of  the  literature,  however,  does  not  substantiate  the 
theory  that  ultra-violet  radiation  is  of  importance  in  air  conditioning, 
since  "the  use  of  ultra-violet  sources  fails  to  produce  indoors,  the  pre- 
viously mentioned  stimulating  qualities  found  in  outdoor  air. 

Experiments35  show  that  in  occupied  rooms  there  is  a  marked  decrease 
in  both  positive  and  negative  small  ions.  As  shown  in  Fig.  5,  soon  after 
the  occupants  assembled  the  ionic  content  fell  abruptly  to  a  very  low  level 
which  was  maintained  until  the  occupants  left  the  room.  Both  positive 
and  negative  ions  began  to  rise  again  as  soon  as  the  occupants  departed. 

The  effects  of  the  decrease  in  the  ionic  content  of  indoor  air  on  comfort 
and  health  have  not  yet  been  subjected  to  sufficient  scientific  investiga- 
tion. It  would  appear,  however,  from  the  evidence  at  hand,  that  comfort 
is  not  associated  with  a  high  ion  content — but  this  must  be  considered,  at 
least  for  the  time  being,  as  still  a  subject  for  further  study. 

»A.S.H.V.E.  research  paper  entitled  Changes  in  Ionic  Content  in  Occupied  Rooms,  Ventilated  by 
Natural  and  Mechanical  Methods,  by  C.  P.  Yaglou,  L.  C.  Benjamin  and  S.  P.  Choate  (A.S.H.V.E.  TRANS- 
ACTIONS, Vol.  37,  1931).  Physiologic  Changes  During  Exposure  to  Ionized  Air,  by  C.  P.  Yaglou,  A.  D. 
Brandt  and  L.  C.  Benjamin  (A.S.H.V.E.  Journal  Section,  Heating,  Piping  and  Air  Conditioning,  August, 
1933).  Diurnal  and  Seasonal  Variations  in  the  Small  Ion  Content  of  Outdoor  and  Indoor  Air,  by  C.  P. 
Yaglou  and  L.  C.  Benjamin,  (A.S.H.V.E.  Journal  Section,  Heating,  Piping  and  Air  Conditioning.  January. 
1934).  The  Nature  of  Ions  in  Air  and  their  Possible  Physiological  Effects,  L.  B.  Loeb  (A.S.H.V.E.  Journal 
Section,  Heating,  Piping  and  Air  Conditioning,  October,  1934). 

56 


CHAPTER  2 — VENTILATION  AND  AIR  CONDITIONING  STANDARDS 


HEAT  AND  MOISTURE  LOSSES 

In  order  to  solve  air  conditioning  problems  involving  the  human  body 
it  is  necessary  to  know  the  rate  at  which  sensible  and  latent  heat  are  given 
up  by  the  body  under  various  conditions  of  temperatue  and  activity. 
Research  at  the  A.S.H.V.E.  Laboratory M»  35  has  resulted  in  the  data  given 
in  Figs,  7,  8,  and  9.  Table  3  gives  the  metabolic  rates  for  various  degrees 
of  activity. 

The  experimental  data  from  which  the  curves  were  drawn  show  that 


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EFFECTIVE    TEMPERATURE     "FAHR. 

FIG.  6.    RELATION  BETWEEN  TOTAL  HEAT  Loss  FROM 

THE   HUMAN   BODY  AND   EFFECTIVE    TEMPERATURE 

FOR  STILL  AiRa 

aCurve  A — Men  working  66,160  ft-lb  per  hour.  Curve  B — 
Men  working  33,075  ft-lb  per  hour.  Curve  C — Men  working 
16,538  ft-lb  per  hour.  Curve  D — Men  seated  at  rest.  Curves  A  and 
C  drawn  from  data  at  an  effective  temperature  of  70  deg  only  and 
extrapolating  the  relation  between  curves  B  and  D,  which  were 
drawn  from  data  at  many  temperatures. 

total  heat  loss  does  not  vary  appreciably  within  the  comfort  zone  (see 
Fig.  6).  Above  or  below  this  range  the  variation  is  approximately  a 
function  of  effective  temperature.  Sensible  and  latent  heat  losses  (Figs. 
7  and ^9)  on  the  other  hand,  vary  greatly  within  the  comfort  zone,  the 
variation  following  closely  the  dry-bulb  temperature. 

Although  total  heat  loss  and  sensible  and  latent  heat  losses  are  not 
exact  functions  of  effective  and  dry-bulb  temperature,  respectively,  for  all 

t?  Jf1^6"11?1  Exchanges  Between  the  Bodies  of  Men  Working  and  the  Atmospheric  Environment,  by 
N  2  M  hCIlVn  W"  Teagne*  W"  E'  Maier'  and  w-  p-  Yant  (American  Journal  of  Hygiene,  Vol.  XIII, 

57 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


conditions  of  humidity  and  air  motion,  they  are  plotted  as  such  in  the 
curves.  This  is  accomplished  by  approximations  which  are  sufficiently 
accurate  for  application  to  practical  problems.  Comparison  of  Figs. 
7  and  8  shows  how  the  cooling  load  may  vary  between  sensible  and  latent 
heat  elimination  for  different  atmospheric  conditions  and  activities  of 
occupants. 

An  atmospheric  condition  resulting  in  sensible  perspiration  is  to  be 


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FIG.  7.     RELATION  BETWEEN  SENSIBLE  HEAT  Loss 

FROM  THE  HUMAN  BODY  AND  DRY-BULB  TEMPERATURE 

FOR  STILL  AiRa 

aCurve  A — Men  working  66,150  ft-lb  per  hour.  Curve  B — 
Men  working  33,075  ft-lb  per  hour.  Curve  C— -Men  working 
16,538  ft-lb  per  hour.  Curve  D — Men  seated  at  rest.  Curves  A  and 
C  drawn  from  data  at  a  dry-bulb  temperature  of  81.3  F  only  and 
extrapolating  the  relation  between  curves  B  and  D  which  were 
drawn  from  data  at  many  temperatures. 

avoided  for  obvious  reasons.  Tables  4  and  5  give  the  approximate  effec- 
tive temperatures  at  which  perspiration  is  noticeable  in  different  degrees 
for  95  per  cent  and  20  per  cent  relative  humidity. 

In  theaters,  auditoriums,  department  stores  and  other  crowded  en- 
closures, the  amount  of  heat  and  moisture  given  off  by  the  people  is  so 
large  that  normal  changes  in  outside  temperature  and  humidity  have 
relatively  little  effect  on  indoor  air  conditions.  The  principal  object  of  air 
conditioning  in  such  places  is  to  remove  excessive  heat  and  moisture  by 
supplying  a  sufficient  quantity  of  properly  conditioned  air.  The  indoor 
air  conditions,  however,  must  be  varied  according  to  the  outside  tem- 
perature, as  has  been  pointed  out. 

58 


CHAPTER  2  —  VENTILATION  AND  Am  CONDITIONING  STANDARDS 


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FIG.  8.    LATENT  HEAT  AND  MOISTURE  Loss  FROM  THE  HUMAN  BODY  BY  EVAPORATION, 
IN  RELATION  TO  DRY-BULB  TEMPERATURE  FOR  STILL  AIR  CONDITIONS^- 

aCmrve  A — Men  working  66,150  ft-lb  per  hour.  Curve  B — Men  working  33,075  ft-Ib  per  hour.'  Carve 
C— Men  working  16,538  ft-Ib  per  hour.  Curve  D — Men  seated  at  rest.  Curves  A  and  C  drawn  from  data 
at  a  dry-bulb  temperature  of  81. 3  F  only  and  extrapolating  the  relation  between  Curves  Band  D  which 
were  drawn  from  data  at  many  temperatures. 


x> 

° 


95° 


FIG.  9.    HEAT  Loss  FROM  THE  HUMAN  BODY  BY  EVAPORATION,  RADIATION  AND  CON- 
VECTION IN  RELATION  TO  DRY-BULB  TEMPERATURE  FOR  STILL  AIR 


aCurve  A  —  Men  working  66,150  ft4f>  per  hoar.  Curve  B—  Men  working  33,075  ft-Ib  per  Lour.  Curve 
C—  Men  working  16,538  ft-lb  Vex  hour.  Curve  D  —  Men  seated  at  rest.  Curves  A  and  C  drawn  from  data 
at  a  dry-bulb  temperature  of  81.3  F  only  and  extrapolating  the  relation  between  Curves  B  and  D  which  were 
drawn  from  data  at  many  temperatures. 

59 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

Although  heat  and  moisture  from  the  human  body  constitute  the  major 
portion  of  the  cooling  load,  in  most  cases  where  air  conditioning  is  pro- 
vided for  comfort  and  health  other  factors  must  also  be  considered.  These 
include  heat  from  lights,  machinery,  and  processes,  as  well  as  the  trans- 
mission and  infiltration  of  heat  through  the  building  structure.  The 
computations  for  these  factors  may  be  made  in  accordance  with  data 
given  in  Chapters  5  and  7. 


TABLE  4. 


CONDITION  OF  SENSIBLE  PERSPIRATION  FOR  PERSONS  SEATED  AT  REST 
UNDER  VARIOUS  ATMOSPHERIC  CONDITIONS* 


DEGBEE  OF  PERSPIRATION* 

95  Per  Cent  Relative 
Humidity 

20  Per  Cent  Relative 
Humidity 

E.  T. 

D.B. 

W.  B. 

E.  T. 

D.B 

W.  B. 

Forehead  clammy                      -  — 

73.0 
73.0 
79.0 
80.0 
84.5 
88.0 
88.5 

73.6 
73.6 
79.7 
80.8 
85.4 
89.0 
89.5 

72.4 
72.4 
78.4 
79.4 
84.0 
87.6 
88.1 

75.0 
75.0 
81.0 
87.0 
86.5 
94.0 
90.0 

87.0 
87.0 
97.5 
109.4 
108.5 
125.2 
116.0 

60.7 
60.7 
67.5 
75.2 
74.6 
85.4 
79.5 

Body  clammy     

Body  damp                   .      .--   ..  —  

Beads  on  forehead 

Body  wet                            

Perspiration  on  forehead  runs  and  drips  
Perspiration  runs  down  body  

ATMOSPHERIC  CONDITION 


aForty  per  cent  of  subjects  registered  degree  of  perspiration  equal  to  or  greater  than  indicated. 


TABLE  5. 


CONDITION  OF  SENSIBLE  PERSPIRATION  FOR  PERSONS  AT  WORK 
UNDER  VARIOUS  ATMOSPHERIC  CONDITIONS^ 


DEGREE  OP  PERSPIRATION* 

95  Per  Cent  Relative 
Humidity 

20  Per  Cent  Relative 

E.T. 

D.B. 

W.B. 

E.Y. 

D.B. 

W.B. 

Forehead  clammy  
Body  clammy  _.  

59.0 
50.0 
60.0 
68.0 
69.0 
78.5 
79.0 

59.4 
50.2 
60.3 
68.5 
69.6 
79.3 
79.8 

58.3 
49.3 
59.3 
67.5 
68.5 
78.0 
78.5 

69.5 
57.0 
62.5 
76.0 
71.0 
82.0 
81.0 

80.5 
61.6 
69.6 
91.0 
82.8 
100.5 
99.8 

56.5 
44.2 
49.5 
63.4 
53.0 
70.2 
69.0 

Body  damp  .  
Beads  on  forehead,  „    

Body  wet                         .          

Perspiration  on  forehead  runs  and  drips  
Perspiration  runs  down  body  

ATMOSPHERIC  CONDITION 


•Forty  per  cent  of  subjects  registered  degree  of  perspiration  equal  to  or  greater  than  indicated. 

In  many  cases,  allowance  must  also  be  made  for  sun  effect  and  for  heat 
capacity  of  the  building  structure  in  accordance  with  studies  by  the 
A.S.H.V.E.  Research  Laboratory37.  Another  item  to  be  considered  is  the 
radiant  heat  received  by  the  body  from  high  temperature  wall  and  celling 
surfaces. 


^Heat  Transmission  as  Influenced  by  Heat  Capacity  and  Solar  Radiation,  by  F.  C.  Houghten,  J.  L.> 
Blackshaw,  E.  M.  Pugh,  and  Paul  McDermott  (A.S.H.V.E.  TRANSACTIONS,  Vol.  38,  1932). 


60 


CHAPTER  2 — VENTILATION  AND  AIR  CONDITIONING  STANDARDS 

PROBLEMS  IN  PRACTICE 

1  •  What  is  the  purpose  and  method  of  conditioning  the  air  of  occupied  rooms? 

Chiefly  comfort,  and  the  method  is  to  control  the  temperature,  humidity,  and  air  distri- 
bution, and  to  prevent  the  accumulation  of  excessive  body  odors  in  the  air.  Other 
factors  have  yet  to  be  studied. 

2  •  What  are  the  most  comfortahle  air  conditions? 

Comfort  standards  are  not  absolute,  but  they  are  greatly  affected  by  the  physical  con- 
dition of  the  individual,  and  the  climate,  season,  age,  sex,  clothing,  and  physical  activity. 
For  the  northeastern  climate  of  the  United  States,  the  conditions  which  meet  the  require- 
ments of  the  majority  of  people  consist  of  temperatures  between  68  and  72  F  in  winter 
and  between  70  and  85  F  in  summer,  the  latter  depending  largely  upon  the  prevailing 
outdoor  temperature.  The  most  desirable  relative  humidity  range  seems  to  be  between 
30  and  60  per  cent. 

3  •  Are  the  optimum  conditions  for  comfort  identical  with  those  for  health? 

There  are  no  absolute  criteria  of  the  prolonged  effects  of  various  air  conditions  on  health. 
For  the  present  it  can  be  only  inferred  that  bodily  discomfort  may  be  an  indication  of 
conditions  that  may  produce  poor  health. 

4  •  Given  dry -bulb  and  wet-bulb  temperatures  of  76  F  and  62  F,  respectively, 
and  an  air  velocity  of  100  fpm,  determine:    (1)  effective  temperature  of  the  con- 
dition; (2)  effective  temperature  with  still  air;  (3)  cooling  produced  by  the  move- 
ment of  the  air;  (4)  velocity  necessary  to  reduce  the  condition  to  66  deg  effective 
tempera  tur  e . 

(lj  In  Fig.  1  draw  line  AB  through  given  dry-  and  wet-bulb  temperatures.  Its 
intersection  with  the  100-ft  velocity  curve  gives  69  deg  for  the  effective  temperature  of 
the  condition .  (2)  Follow  line  A  B  to  the  right  to  its  intersection  with  the  20-f pm  velocity 
line,  and  read  70.4  deg  for  the  effective  temperature  for  this  velocity  or  so-called  still  air. 
(3)  The  cooling  produced  by  the  movement  of  the  air  is  70.4  —  69  =  1.4  deg  effective 
temperature.  (4)  Follow  line  AB  to  the  left  until  it  crosses  the  66  deg  effective  tempera- 
ture line  and  interpolate  velocity  value  of  340  fpm  to  which  the  movement  of  the  air 
must  be  increased. 

5  •  Given  dry-bulb  and  wet-bulb  temperatures  of  75  and  68  F,  respectively, 
first,  what  is  the  effective  temperature?    Second,  is  this  condition  warmer  or 
cooler  than  80  F  dry-bulb  and  60  F  wet-bulb? 

The  first  condition  is  given  by  the  intersection  of  the  75  F  dry-bulb  line  and  the  68  F  wet- 
bulb  line  (Fig.  3).  The  effective  temperature  of  72.1  deg  is  given  by  the  numerical  value 
of  the  effective  temperature  line  passing  through  this  point  and  indicated  by  the  scale 
along  the  saturation  curve.  The  second  condition  is  given  by  the  intersection  of  80  F 
dry-bulb  and  60  F  wet-bulb  and  is  71.8  deg  ET.  It  is  therefore  0.3  deg  ET  cooler  than 
the  first  condition. 

6  •  Given  76  F  dry-bulb  and  61  F  wet-bulb,  how  many  degrees  difference 
are  there  between  this  condition  and  the  winter  comfort  line  or  66  deg  ET? 

The  effective  temperature  for  this  condition  is  given  by  the  intersection  of  the  76-F  dry- 
bulb  and  61-F  wet-bulb  lines  and  is  70  deg  ET,  which  is  4  deg  ET  warmer  than  the 
comfort  line. 

7  •  Assume  that  the  design  of  an  air  conditioning  system  for  a  theater  is  to  be 
based  on  an  outdoor  dry-bulb  temperature  of  95  F  and  a  wet-bulb  temperature 
of  78  F  with  an  indoor  relative  humidity  of  50  per  cent.    According  to  Table  2, 
the  dry-bulb  temperature  in  the  auditorium  should  be  80  F.     Estimate  the 
sensible  and  latent  heat  given  up  per  person. 

The  sensible  heat  given  up  per  person  per  hour  under  this  condition  may  be  obtained 
from  Fig.  7.  With  an  abscissa  value  of  80  F,  Curve  D  for  men  seated  at  rest  gives  a  value 
on  the  ordinate  scale  of  220  Btu  per  person  per  hour  as  the  sensible  heat  loss.  The  latent 
heat  given  up  by  a  person  seated  at  rest  per  hour  may  be  obtained  from  Fig.  8.  With  an 

61 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

abscissa  value  of  SO  FT  Curve  D  indicates  a  latent  heat  loss  of  175  Btu  per  hour  left  hand 
scale)  or  a  moisture  loss  of  1190  grains  per  hour  (right  hand  scale). 

8  •  How  much  sensible  heat,  how  much  latent  heat  and  how  much  water 
vapor  wUl  be  added  per  hour  to  the  atmosphere  of  an  auditorium  by  an  audience 
of  1000  adults,  when  the  dry-  and  wet-bulb  temperatures  are  75  F  and  63.5  F, 
respectively? 

From  Curve  D,  Fig,  7,  find  the  sensible  heat  loss  per  person  for  a  dry-bulb  temperature 
of  75  F  and  still  air  to  be  265  Btu  per  hour.  From  Fig.  8  find  the  latent  heat  loss  per 
person  for  a  dry-bulb  temperature  of  75  F  to  be  134  Btu  per  hour  and  the  moisture 
added  to  be  905  grains  per  hour.  Sensible  heat  =  1000  X  265  =  265,000  Btu.  Latent 
heat  =  1000  X  134  =  134,000  Btu.  Water  vapor  added  per  hour  to  the  air  in  the 
auditorium  =  1000  X  905  =  905,000  grains  or  129  Ib. 

The  sensible  and  latent  heat  added  to  the  air  may  also  be  found  as  follows:  The  effective 
temperature  for  dry-  and  wet-bulb  temperatures  of  75  F  and  63.5  F,  respectively,  is 

70.3  deg.    From  Curve  D,  Fig.  6,  find  403  Btu  as  the  total  heat  added  to  the  air  by  a 
person  for  an  effective  temperature  of  70.3  deg.     From  Fig.  9  find  the  percentage  of 
sensible  and  latent  heat  at  a  dry-bulb  temperature  of  75  F  to  be  66.5  per  cent  and  33.5 
per  cent.    The  sensible  heat  added  to  the  air  in  the  auditorium  is  1000  X  0.665  X  403  = 
267,995  Btu  per  hour.    The  latent  heat  added   is  1000  X  0.335  X  403  =  135,005  Btu 
per  hour. 

9  •  If  the  dry-  and  wet-bulb  temperatures  of  the  auditorium  were  85  F  and 
63  F,  respectively,  how  much  heat  and  moisture  would  be  dissipated  to  the 
atmosphere? 

From  Figs.  7  and  8,  respectively,  the  sensible  and  latent  heat  losses  per  person  for  a  dry- 
bulb  temperature  of  85  F  are  found  to  be  164  and  225  Btu  per  hour.  The  water  vapor 
added  to  the  atmosphere  is  1520  grains  per  hour.  The  audience  will  then  add  164,000 
Btu  sensible  heat,  225,000  Btu  latent  heat  and  1,520,000  grains  or  217  Ib  of  water  vapor 
to  the  air  in  the  auditorium  per  hour. 

10  •  Neglecting  the  gain  or  loss  of  heat  to  an  auditorium  by  transmission  or 
infiltration  through  the  walls,  windows  and  doors,  how  many  cubic  feet  of 
outside  air,  with  dry-  and  wet-bulb  temperatures  of  65  F  and  59  F,  respectively, 
(63.1  deg  ET)  must  be  supplied  per  hour  to  an  auditorium  containing  1000 
people  in  order  that  the  inside  shall  not  exceed  75  F  (dry-bulb)  and  65  F  (wet- 
bulb),  respectively? 

Figs.  7  and  8  give  265  Btu  sensible  heat  and  905  grains  of  moisture  as  the  additions  per 
person  with  a  dry-bulb  temperature  of  75  F  in  the  auditorium.  Therefore,  265,000  Btu 
of  sensible  heat  and  905,000  grains  of  moisture  will-be  added  to  the  air  in  the  auditorium 
per  hour. 
Taking  0.24  as  the  specific  heat  of  airf  2.4  Btu  per  pound  of  air  will  be  required  to  raise 

oopr  QAQ 

the  dry~bulb  temperature  from  65  to  75  F  and         '        =  110,400  Ib  of  air  or  110,400  X 

^.4 

1  47Q  000 

13.4  =  1,479,000  cfh  of  air  will  be  required.    This  is  equivalent  to  ^A  J/gn  =  24-7  cfm 

ILKJU  X  oU 
per  person. 

The  moisture  content  of  the  inside  air  as  taken  from  a  psychrometric  chart  is  76  grains 
per  pound  of  dry  air  and  that  of  the  outside  condition  is  65  grains.  The  increase  in 

905  000 
moisture  content  will  therefore  be  11  grains  per  pound  of  dry  air.     Hence — ~~^- —  = 

82,300  Ib  of  air  at  the  specified  condition  will  be  required.    This  is  equivalent  to  82,300 

i  1 02  OOO 
X  13.4  =  1,103,000  cfh  of  air  or  /QQQ  ^  6Q  «  18.4  cfm  of  air  per  person. 

The  higher  volume  of  24.7  cfm  per  person  will  be  required  to  keep  the  dry-bulb  tem- 
perature from  rising  above  the  75  F  specified.  The  wet-bulb  temperature  will  therefore 
not  rise  to  the  maximum  of  65  F. 

11  •  Assume  that  a  man  performs  work  at  a  rate  equivalent  to  50,000  ft-lb  per 
hour,  in  an  atmosphere  having  a  dry-bulb  temperature  of  70  F.    Estimate  the 
sensible  and  latent  heat  given  off  per  hour. 

62 


CHAPTER  2 — VENTILATION  AND  AIR  CONDITIONING  STANDARDS 

Since  the  net  mechanical  efficiency  of  the  human  body  is  about  20  per  cent,  the  increase 

oO  000 

in  metabolism  due  to  work,  over  the  resting  metabolism,  will  be  -^,--Lrn-oh  =  ^20  Btu 

/  /  o    X  U.^U 

per  hour.  Assuming  a  resting  metabolism  of  400  Btu  per  hour  ^see  Fig.  6),  the  total 
metabolism  during  work  will  be  400  -f-  320  =  720  Btu  per  hour,  and  the  total  heat  loss 

720  —  ^  _',    -  =  656  Btu  per  hour,  approximately.    In  Fig.  9,  follow  a  vertical  line  from 

/  /o 

a  dry-bulb  temperature  of  70  F  to  a  point  midway  between  Curves /I  and  B.  The  sensible 
heat  loss  is  about  46  per  cent  of  the  total  loss,  or  0.46  X  656  =  302  Btu  per  hour,  and  the 
latent  heat  is  54  per  cent  of  the  total  or  0.54  X  656  =  354  Btu  per  hour. 

12  •  The  characteristics  of  air  supplied  to  ventilate  a  room  are: 

Carbon  dioxide  concentration    ,         .   .   .       A  parts  per  10,000 

Wet-bulb  temperature .  -  45.2  F 

Dry-bulb  temperature ,  . .   .   „     55.0  F 

Moisture  content 29.0  grains  per  pound  of  dry  air 

a.  What  will  be  the  dry -bulb  temperature  of  the  air  in  the  room  if  it  is  occupied 
by  five  adults,  if  the  air  change,  including  both  ventilation  and  infiltration,  is 
50  cu  ft  per  minute,  and  assuming  that  there  is  no  heat  gain  or  loss  to  the  room 
from  any  source  other  than  from  the  occupants? 

b.  What  will  be  the  carbon  dioxide  concentration  of  the  air  in  the  room  under 
these  conditions? 

c.  What  will  be  the  moisture  content  of  the  ah*  in  the  room  under  these  con- 
ditions? 

d.  What  will  be  the  wet-bulb  temperature  and  the  relative  humidity  of  the  air 
in  the  room  under  these  conditions? 

e.  WTiat  would  the  temperature  of  the  incoming  air  have  to  be  to  give  a  room  a 
dry-bulb  temperature  of  70  F? 

a.  The  air  change  is  10  cu  ft  per  minute  per  occupant.    From  the  bottom  chart  of  Fig.  4 
at  the  intersection  of  an  incoming  air  dry-bulb  temperature  of  55.0  F  and  a  rate  of  air 
supply  of  10  cu  ft  per  minute  per  occupant,  find  by  interpolation  between  the  70  F  and 
80  F  adult  curves  the  dry-bulb  temperature  of  the  air  in  the  room  to  be  78.0  F. 

b.  From  the  top  chart  of  Fig.  4  find  the  increase  in  CO2  concentration  to  be  10  parts  of 
CC>2  per  10,000  parts  of  air.    Therefore,  the  air  in  the  occupied  room  will  contain  14 
parts  of  CO2  per  10,000. 

c.  From  the  center  chart  in  Fig,  4  find  by  interpolation  between  the  70  F  and  80  F  adult 
curves  the  increase  in  moisture  content  to  be  23  grains  per  pound  of  dry  air  for  adults  in 
78  F  air.    This  gives  a  resultant  moisture  content  of  the  air  in  the  room  of  52  grains  per 
pound  of  dry  air. 

d.  From  the  psychrometric  chart,  Fig.  3,  find  the  resulting  wet-bulb  temperature  and 
relative  humidity  for  78  F  dry-bulb  and  52. grains  of  moisture  to  be  61.0  F  and  37  per 
cent,  respectively. 

e.  From  the  bottom  chart,  Fig.  4,  find  the  required  incoming  air  temperature  to  be  42  F 
dry-bulb. 

13  •  Name  three  factors  that  influence  the  feeling  of  warmth  and  the  elimina- 
tion of  body  heat. 

Temperature,  humidity,  and  air  movement. 

14  •  What  is  meant  by  effective  temperature? 

Effective  temperature  is  a  composite  index  which  combines  the  measurements  of  tem- 
perature, air  motion,  and  humidity  into  a  single  value.  It  is  an  arbitrary  index  of  the 
degree  of  warmth  or  cold  felt  by  the  human  body  due  to  these  factors. 

15  •  Referring  to  the  A.S.H.V.E.  Comfort  Chart  (Fig.  3),  list  the  conditions 
(dry-bulb,  wet-bulb,  effective  temperature,  and  humidity)  which  will  produce 
comfort  at  each  corner  of  the  average  winter  comfort  zone  and  of  the  average 
summer  comfort  zone. 


*Heat  equivalent  of  raechaiacal  warfc  IB  foat^poHiwis  per  Btu, 

63 


AMERICAS  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 
A\erage  winter  comfort  zone: 


WET-BULB 

DRT-BULB 

RELATIVE 

F 

F 

HUMIDITY 

58.5 

64.5 

70  per  cent 

51.5 

67.5 

30  per  cent 

59.0 

79.0 

30  per  cent 

67.0 

74.0 

70  per  cent 

Average  summer  comfort  zone: 


WET-BULB 

DRY-BULB 

RELATIVE 

F 

F 

HUMIDITY 

62.0 

68.0                   ! 

70  per  cent 

54.0 

72.0 

30  per  cent 

63.5 

:                85.0 

30  per  cent 

71.5 

78.5 

70  per  cent 

16  •  What  is  generally  considered  to  be  the  desirable  and  practicable  range  of 
relative  humidity  indoors? 

30  per  cent  to  60  per  cent. 


64 


Chapter  3 

INDUSTRIAL  AIR  CONDITIONING 

Moisture   Content   and   Regain,   Hygroscopic  Materials,   Atmos- 
pheric Conditions  Required?  Air  Conditioning  of  Libraries,  Banana 
Ripening,   Lumber  Drying,    Greenhouse  Heating,   Apparatus  for 
Industrial  Conditioning 

AIR  conditioning  is  applicable  to  industrial  or  process  conditioning  for 
the  improvement  of  products  during  manufacture,  or  for  making  the 
process  independent  of  climatic  conditions.  In  many  industries,  the 
temperature  and  relative  humidity  of  the  air  have  a  marked  influence  upon 
the  rate  of  production  and  the  weight,  strength,  appearance,  and  general 
quality  of  the  product.  These  results  are  due  to  the  fact  that  most 
materials  of  animal  or  vegetable  origin,  and  to  a  lesser  extent  minerals  in 
certain  forms,  either  take  up  or  give  moisture  to  the  surrounding  air. 

MOISTURE  CONTENT  AND  REGAIN 

The  terms  moisture  content  and  regain  refer  to  the  amount  of  moisture 
in  hygroscopic  materials.  Moisture  content  is  the  more  general  term  and 
refers  either  to  free  moisture  (as  in  a  sponge)  or  to  hygroscopic  moisture 
(which  varies  with  atmospheric  conditions) .  It  is  usually  expressed  as  a 
percentage  of  the  total  weight  of  material.  Regain  is  more  specific  and 
refers  only  to  hygroscopic  moisture.  It  is  expressed  as  a  percentage  of  the 
bone-dry  weight  of  material.  For  example,  if  a  sample  of  cloth  weighing 
100.0  grains  is  dried  to  a  constant  weight  of  93.0  grains,  the  loss  in  weight, 
or  7.0  grains,  represents  the  weight  of  moisture  originally  contained.  This 
expressed  as  a  percentage  of  the  total  weight  (100.0  grains)  gives  the 
moisture  content  or  7  per  cent.  The  regain,  which  is  expressed  as  a  per- 

7.0 

centage  of  the  bone-dry  weight,  is     '  A  or  7.5  per  cent. 

yo.u 

The  use  of  the  term  regain  does  not  necessarily  imply  that  the  material 
as  a  whole  has  been  completely  dried  out  and  has  re-absorbed  moisture. 
In  the  case  of  certain  textiles,  for  instance,  complete  drying  during  manu- 
facturing is  avoided  as  it  might  appreciably  reduce  the  ability  of  the 
material  to  re-absorb  moisture.  In  measuring  moisture  it  is  necessary 
to  dry  out  a  sample  so  that  the  loss  in  weight  may  be  used  as  a  basis  for 
calculating  the  regain  of  the  whole  lot. 

65 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


TABLE  1.    REGAIN  OF  HYGROSCOPIC  MATERIALS 

Moisture  Content  Expressed  in  Per  Cent  of  Dry  Weight  of  the  Substance  at 
Various  Relative  Humidities — Temperature,   75   F 


CLASSI- 
FICATION 

MATERIAL 

DESCRIPTION 

RELATIVE  HUMIDITY—  PER  CENT 

AUTHORITY 

10 

20 

30 

40 

5.5 

50 

60 

70 

80 

90 

Natural 
Textile 
Fibres 

Cotton 

Sea  island  —  roving 

2.5 

3.7 

4.6 

6.6 

7.9 

9.5 

11.5 

14.1 

Hartshorne 

|  Cotton 

American-  cloth 

2.6 

3.7 

4.4 

5.2 

5.9 

6.8 

8.1 
22.8 

10.0 

14.3 

Schloesing 

Cotton 

Absorbent 

4.8 

9.0 

12.5 

15.7 

18.5 

20.8 

24.3 

25.8 

Fuwa 

Woo! 

Australian  merino  —  skein 

4.7 

7.0 

8.9 

10.8 

12.8 

14.9 

17.2 

19.9 

23.4 

Hartshorne 

Silk 

Raw  ehevennes  —  skein 

3.2 

5.5 

6.9 

8.0 

8.9 

10.2 

11.9 

14.3 

18.8 

Schloesing 

Linen 

Table  cloth 

1.9 

2.9 

3.6 

4.3 

5.1 

6.1 

7.0 

8.4 

10.2 

Atkinson 

Linen 

Dry  spun—  yarn 

3.6 

5.4 

6.5 

7.3 

8.1 

8.9 

9.8 

11.2 

13.8 

Sommer 

Jute 

Average  of  several  grades 

3.1 

5.2 

6.9 

8.5 

10.2 

12.2 

14.4 

17.1 

20.2 

Storch 

Hemp 

Manila  and  sisal—  rope 

2.7 

4.7 

6.0 

7.2 
7.9 

8.5 

9.9 
10.8 

11.6 

13.6 

15.7 

Fuwa 

Bayona 

Viscose  Nitrocellu- 
lose Cupramonium 

Average  skein 

4.0 

5.7 

6.8 

9.2 

12.4 

14.2 

16.0 

Robertson 

Cellulose  Acetate 

Fibre 

0.8 

1.1 

1.4 

1.9 

4.7 

2.4 

3.0 
6.1 

3.6 

4.3 
8.7 

5.3 

Robertson 

Paper 

M.  F.  Newsprint 

Wood  pulp  —  24%  ash 

2.1 

3.2 

4.0 

5.3 

7.2 

10.6 

U.  S.  B.  of  S. 

H,  M.  F.  Writing 

Wood  pulp  —  3%  ash 

3.0 

4.2 

5.2 

6.2 

7.2 

8.3 

9.9 

11.9 

14.2 
13.2 

U.S.B.ofS. 

White  Bond 

Rag—  1%  ash 

2.4 

3.7 

4.7 

5.5 

6.5 
6.2 

7.5 

8.8 

10.8 

U.S.  B.  ofS. 

Com.  Ledger 

75%  rag—  1%  ash 

3.2 

4.2 
4.6 

5.0 

5.6 

6.9 

8.1 

10.3 

13.9 

U.S.B.ofS. 

Kraft  Wrapping 

Coniferous 

3.2 

5.7 

6.6 

7.6 

8.9 

10.5 

12.6 

14.9 

U.S.B.ofS. 

Misc. 
Organic 

Materials 

Leathefr 

Sole  oak—  tanned 

5.0 

8.5 

11.2 

13.6 

16.0 

18.3 

20.6 

24.0 

29.2 

Phelps 

Catgut 

Racquet  strings 

4.6 

7.2 

8.6 

10.2 
6.6 
0.44 

12.0 
7.6 
0.54 

14.3 

17.3 
10.7 

19.8 

21.7 

Fuwa 

Glue 

Hide 

3.4 
0.11 

4.8 
0.21 

5.8 

9.0 

11.8 

12.5 

Fuwa 

Rubber 

Solid  tire 

0.32 

0.66 

0.76 

0.88 

0.99 

Fuwa 

Wood 

Timber  (average) 

3.0 

4.4 

5.9 

7.6 

9.3 

11.3 

14.0 

17.5 

22.0 

Forest  P.  Lab. 

Soap 

White 

1.9 

3.8 

5.7 

7.6 
13.3 

10.0 

12.9 

16.1 
25.0 

19.8 

23.8 

Fuwa 

Tobacco 

Cigarette 

5.4 

8.6 

11.0 

16.0 

19.5 

33.5 

50.0 

Ford 

Food- 
stuffs 

White  Bread 

0.5 

1.7 

3.1 

4.5 

6.2 

8.5 

11.1 

14.5 

19.0 

Atkinson  - 

Crackers 

2.1 

2.8 

3.3 

3.9 

5.0 

6.5 
13.7 

8.3 

10.9 

14.9 

Atkinson 

Macaroni 

5.1 

7.4 

8.8 

10.2 

11.7 

16.2 

19.0 

22.1 

Atkinson 

Flour 

2.6 
2.2 

4.1 
3.8 

5.3 

6.5 

8.0 

9.9 

12.4 

15.4 

19.1 

Bailey 

Starch 

5.2 

6.4 

7.4 

8.3 

9.2 

10.6 

12.7 

Atkinson 

Gelatin 

0.7 

1.6 

2.8 

3.8 

4.9 

6.1 

7.6 

9.3 

11.4 

Atkinson 

Miac. 
Inorganic 
Materials 

Asbestos  Fibre 

Finely  divided 

0.16 

0.24 

0.26 

0.32 

0.41 

0.51 

0.62 

0.73 

0.84 

Fuwa 

Silica  Gel 

5.7 

9.8 

12.7 

15.2 

17.2 

18.8 

20.2 

21.5 

22.6 

Fuwa 

Domestic  Coke 

0.20 

0.40 

0.61 

0.81 

1.03 

1.24 

1.46 

1.67 

1.89 

Selvig 

Activated  Charcoal 

Steam  activated 

7.1 

14.3 

22.8 

26.2 

28.3 

29.2 

30.0 

31.1 

32.7 

Fawa 

Sulphuric  Acid 

H*SOt 

33.0 

41.0 

47.5 

52.5 

57.0 

61.5 

67.0 

73.5 

82.5 

Mason 

66 


CHAPTER  3 — INDUSTRIAL  AIR  CONDITIONING 


HYGROSCOPIC  MATERIALS 

Air  conditioning  is  extensively  used  in  the  manufacture  or  processing  of 
hygroscopic  materials  such  as  textiles,  paper,  wood,  leather,  tobacco,  and 
foodstuffs.  Where  the  physical  properties  of  the  product  affect  value,  the 
question  of  moisture  is  of  special  importance.  With  increase  in  moisture 
content,  hygroscopic  materials  ordinarily  become  softer  and  more  pliable. 
Economy  of  manufacturing,  therefore,  requires  that  the  moisture  content 
be  maintained  at  a  percentage  most  favorable  to  rapid  and  satisfactory 
manipulation  and  to  a  minimum  loss  of  material  through  breakage.  A 
constant  condition  is  desirable  in  order  that  high  speed  machinery  may  be 
adjusted  permanently  for  the  desired  production  with  a  minimum  loss 
from  delays,  wastage  of  raw  material,  and  defective  product. 

In  the  processing  of  hygroscopic  materials,  it  is  usually  necessary  to 
secure  a  final  moisture  content  suitable  for  the  goods  as  shipped.  Where 
the  goods  are  sold  by  weight  it  is  proper  that  they  contain  a  normal  or 
standard  moisture  content.  Air  conditioning  is  important  in  certain 
branches  of  the  chemical  industry  in  controlling  the  temperature  of 
reaction  and  facilitating  or  retarding  evaporation.  The  control  of 
moisture  content  of  air  supplied  to  blast  furnaces  in  the  manufacture  of 
pig  iron  also  has  proved  advantageous. 

The  moisture  content  of  a  hygroscopic  material  at  any  time  depends 
upon  the  nature  of  the  material  and  upon  the  temperature  and  especially 
the  relative  humidity  of  the  air  to  which  it  has  been  exposed.  Not  only 
do  different  materials  acquire  different  percentages  of  moisture  after 
prolonged  exposure  to  a  given  atmosphere,  but  the  rate  of  absorption  or 
drying  out  varies  with  the  nature  of  the  material,  its  thickness  and 
density. 

Table  1  shows  the  regain  or  hygroscopic  moisture  content  of  several 
organic  and  inorganic  materials  when  in  equilibrium  at  a  dry-bulb  tem- 
perature of  75  F  and  various  relative  humidities.  The  effect  of  relative 
humidity  on  regain  of  hygroscopic  substances  is  clearly  indicated.  The 
effect  of  temperature  is  comparatively  unimportant.  In  the  case  of 
cotton,  for  instance,  an  increase  in  temperature  of  10  deg  has  the  same 
effect  on  regain  as  a  decrease  in  relative  humidity  of  one  per  cent.  Changes 
in  temperature  do,  however,  affect  the  rate  of  absorption  or  drying. 
Sudden  changes  in  temperature  cause  temporary  fluctuations  in  regain 
even  when  the  relative  humidity  remains  stationary. 

Conditioning  and  Drying 

Exposure  of  hygroscopic  materials  to  an  atmosphere  of  controlled 
humidity  and  temperature  for  the  purpose  of  establishing  a  specified 
moisture  condition  in  the  material  is  called  conditioning.  Where  the 
desired  final  moisture  content  is  relatively  low,  the  term  drying  is  usually 
used-  In  any  case,  control  of  relative  humidity,  temperature,  air  velocity 
and  length  of  exposure  are  all  of  more  or  less  importance. 

The  conditioning  treatment  may  be  undertaken  in  a  special  enclosure 
(conditioning  room)  or  it  may  be  accomplished  in  the  same  room  and  at 
the  same  time  as  some  regular  manufacturing  process.  For  instance,  in 
the  weaving  of  textiles  a  high  relative  humidity  is  commonly  employed  to 
keep  the  yarn  strong  and  pHafole*  thus  assisting  in  the  weaving  process  and 

67 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


TABLE  2.    DESIRABLE  TEMPERATURES  AND  HUMIDITIES  FOR  INDUSTRIAL  PROCESSING 


I   TEMPERATURE 
INDUSTRY                                                     PROCESS                                               DEGREES 
FAHRENHEIT     ' 

RELATIVE 
HUMIDITY 

PER  CENT 

AUTOMOBILE  Assembly  line  65 

40 

Cake  icing                                          -  - 

70 

50 

Cake  mixing         

75 

65 

Dough  fermentation  room 

80 

76  to  80 

Loaf  cooling 

70 

60  to  70 

Make-up  room               

75  to  80 

55  to  70 

BAKING  

*VTixinsr  room 

75  to  80 

1     55  to  70 

Paraffin  paper  wrapping  

1          80 

55 

Proof  boxes 

80  to  90 

80  to  95 

Storage  of  flour  

70  to  80 

60 

I  Storage  of  yeast      

28  to  40 

60  to  75 

BIOLOGIC  A.L 

Vaccines                       

'    below  32 

PRODUCTS 

Antitoxins 

38  to  42 

Fermentation  in  vat  room 

44  to  50 

50 

BREWING  

Storasre  of  strains 

60 

QA  to  4.^ 

Drying  of  auger  machine  brick  

180  to  200 

Drying  of  refractory  shapes    

110  to  150 

50  to  60 

CERAMIC.-  

JVtoldinor  room. 

80 

60 

Storage  of  clay      

60 

35 

CHEMICAL 

General  storage  

60  to  80 

35  to  50 

Chewing  gum  rolling 

75 

50 

Chewing  gum  wrapping               

70 

45 

Chocolate  covering  

62  to  65 

50  to  55 

CONFECTIONERY 

Hard  candy  making  

70  to  80 

30  to  50 

Packing 

65 

50 

Starch  room  

75  to  85 

50 

Storage 

60  to  68 

50  to  65 

General  manufacture 

60 

45 

DISTILLERY  

Storage  of  grains 

60 

30  to  45 

DRUG 

Storage  of  powders  and  tablets 

70  to  80 

30  to  35 

Insulation  winding 

104 

5 

Manufacture  of  cotton  covered  wire  

60  to  80 

60  to  70 

ELECTRICAL 

Manufacture  of  electrical  win-dings  
Storage  of  electrical  goods  

60  to  80 
60  to  80 

35  to  50 
35  to  50 

Butter  making  

60 

60 

Dairy  chill  room.  „ 

40 

60 

Preparation  of  cereals  

60  to  70 

38 

Preparation  of  macaroni  

70  to  80 

38 

Ripening  of  meats 

40 

80 

FOOD 

Slicing  of  bacon                  .               

60 

45 

Storage  of  apples 

31  to  34 

75  to  85 

Storage  of  citrus  fruit        .                 

32 

80 

Storage  of  eggs  in  shell 

30 

80 

Storage  of  meats  

Oto  10 

50 

Storage  of  susar 

80 

35 

Drying  of  furs 

110 

FUR  

£-»,                        r   r 

OK  4-f*.   Af\ 

Storage  of  furs      

28  to  40 

Zb  to  40 

68 


CHAPTER  3 — INDUSTRIAL  AIR  CONDITIONING 


TABLE  2.    DESIRABLE  TEMPERATURES  AND  HUMIDITIES  FOR  INDUSTRIAL  PROCESSING 

(Continued) 


ISDUSTRT 


PROCESS 


TEMPERATURE 

DEG  SITES 
FAHBINHEIT 


RELATIVE 
HUMIDITY 


INCUBATORS.. 


Chicken j     99  to  102       55  to  75 


LABORATORY 


General  analytical  and  physical- 
Storage  of  materials 


60  to  70 
60  to  70 


60  to  70 
35  to  50 


LEATHER.-. •   Drying  of  hides.. 


90 


LIBRARY I  Book  storage  (see  discussion  in  thischapter)  i     65  to  70         38  to  50 


LINOLEUM.- !   Printing 


80 


40 


MATCH.. 


Manufacturing 

Storage  of  matches.. 


72  to  74     i 
60 


50 


MUNITIONS Fuse  loading 


70 


Drying  of  lacquers 

60  to  80 

25  to  50 

PAINT 

Drying  of  oil  paints.     „  

60  to  90 

25  to  50 

Brush  and  spray  painting 

60  to  80 

25  to  50 

PAPER- 

Binding,  cutting,  drying,  folding,  gluing.. 

60  to  80 

25  to  50 

Storage  of  paper.  _    

60  to  80 

35  to  45 

Development  of  film  

70  to  75 

60 

Drying      .                    ....                      ... 

75  to  80 

50 

PHOTOGRAPHIC.... 

Printing 

70 

70 

Cutting 

72 

65 

Binding 

70 

45 

Folding  

77 

65 

PRINTING 

Press  room  (general)          . 

75 

60  to  78 

Press  room  (lithographic) 

60  to  75 

20  to  60 

Storage  of  rollers 

60  to  80 

35  to  45 

Manufacturing  

90 

RUBBER 

Dipping  of  surgical  rubber  articles  

75  to  80 

25  to  30 

Standard  laboratory  tests 

80  to  84 

42  to  48 

SOAP 

Drying 

110 

70 

Cotton  —  carding      

75  to  80 

50 

combing  —     .  ..     

75  to  80 

60  to  65 

roving 

75  to  80 

50  to  60 

spinning                          _ 

60  to  80 

60  to  70 

weaving 

68  to  75 

70  to  80 

Rayon  —  spinning 

70 

85 

TEXTILE.. 

twisting  

70 

65 

Silk  —        dressing    .. 

75  to  80 

60  to  65 

spinning 

75  to  80 

65  to  70 

throwing 

75  to  80 

65  to  70 

weaving.  

75  to  80 

60  to  70 

Wool  —     carding     _..  .    .  .              

75  to  80 

65  to  70 

spinning 

75  to  80 

55  to  60 

weaving 

75  to  80 

50  to  55 

Cigar  and  cigarette  making,....  

70  to  75 

55  to  65 

TOBACCO  _  

Softening 

90 

85 

Stemming  or  stripping 

75  to  85 

70 

69 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

at  the  same  time  leaving  the  product  in  a  satisfactory  condition  of  regain 
for  commercial  reasons. 

As  a  rule,  commercial  regain  standards  are  specified  percentages  which 
by  test  have  been  found  equivalent  to  a  so-called  standard  atmosphere 
with  which  the  goods  would  be  in  hygroscopic  equilibrium  after  prolonged 
exposure.  Committee  D13  on  Textiles  of  the  American  Society  for 
Testing  Materials  has  adopted  a  relative  humidity  of  64  to  66  per  cent  and 
a  temperature  of  70  to  80  F  as  the  standard  atmosphere  for  textile  testing. 

ATMOSPHERIC  CONDITIONS  REQUIRED 

The  most  desirable  relative  humidity  during  processing  depends  upon 
the  product  and  the  nature  of  the  process.  As  far  as  the  behavior  of  the 
material  itself  and  its  desired  final  condition  are  concerned,  each  material 
and  process  represents  a  different  problem.  The  best  relative  humidity 
may  range  up  to  100  per  cent.  Similarly  the  most  desirable  temperature 
may  range  between  wide  limits  for  different  materials  and  treatments. 
Extremes  in  either  relative  humidity  or  temperature  require  relatively 
expensive  equipment  for  maintaining  these  conditions  and  controlling 
them  automatically.  Also,  in  departments  where  people  are  working, 
their  health,  comfort,  and  productive  efficiency  must  be  considered.  A 
compromise  often  is  desirable. 

It  is  generally  considered  that  relative  humidities  below  40  per  cent 
are  on  the  dry  side,  conducive  to  low  regains,  a  brittle  condition  of 
fibrous  materials,  prevalence  of  static  electricity,  and  a  tendency  toward 
dryness  of  the  skin  and  membranes  of  human  beings.  At  the  other  end 
of  the  scale,  humidities  above  80  per  cent  are  relatively  damp,  conducive 
to  high  regains,  extreme  softness,  and  pliability. 

Table  2  lists  desirable  temperatures  and  humidities  for  industrial  pro- 
cessing. In  using  this  table,  care  must  be  taken  in  qualifying  the  process. 
In  preparing  many  materials,  conditions  are  not  maintained  constantly, 
but  different  temperatures  and  humidities  are  held  for  varying  lengths  of 
time. 

AIR  CONDITIONING  OF  LIBRARIES1 

Temperature  has  little  effect  on  the  preservation  of  books.  A  tempera- 
ture over  100  F,  combined  with  low  relative  humidity,  may  cause  the  book 
materials  to  become  brittle,  while  a  temperature  much  below  freezing  may 
cause  permanent  deterioration  "of  the  glue  in  the  binding.  The  relative 
humidity  should  be  maintained  between  40  and  70  per  cent,  although 
these  limits  need  not  hold  for  short  periods  of  time.  If  the  relative 
humidity  gets  much  below  40  per  cent,  first  the  glue  and  then  the  paper 
will  tend  to  become  brittle  which  will  not  cause  any  permanent  damage 
unless  the  book  is  used  while  in  this  condition,  as  a  subsequent  increase 
in  humidity  will  bring  the  materials  back  to  their  normal  condition.  If 
the  relative  humidity  gets  above  80  per  cent,  the  growth  of  mildew  may 
be  expected. 

One  of  the  principal  agents  of  destruction  and  deterioration  of  paper 
and  books  in  libraries  is  sulphur  dioxide  gas  in  the  air.  If  air  containing 

iSec  U.  S.  Bureau  of  Standards  Bulletin  No.  128  entitled  A  Survey  of  Storage  Conditions  in  Libraries, 
by  Kimberly  and  Hicks. 

70 


CHAPTER  3 — INDUSTRIAL  AIR  CONDITIONING 


sulphur  dioxide  is  allowed  to  come  in  contact  with  cellulose,  the  principal 
constituent  of  paper,  sulphuric  acid  is  formed  on  the  surface.  This  acid 
is  not  volatile  at  ordinary  temperatures  and  therefore  accumulates 
throughout  the  life  of  the  paper.  The  destructive  effect  of  the  acid  on  the 
paper  is  independent  of  the  relative  humidity  of  the  surrounding  air. 
Low  alkaline  concentration  spray  water  may  be  used  in  an  air  washer  to 
neutralize  the  acid  condition.  Such  an  air  washer  must  be  especially 
constructed  to  resist  corrosion. 

BANANA  RIPENING 

Ripe  bananas  are  very  perishable  and  for  this  reason  men  who  deal  in 
them  must  depend  mainly  upon  control  of  the  ripening  speed  as  a  means 
of  regulating  their  daily  supply  of  the  fruit.  Knowledge  and  experience 
are  required  in  regulating  the  ripening  treatment  and  to  control  the 
ripening  speed.  An  accurate  appraisal  must  be  based  upon  a  careful 
examination  of  the  fruit  when  received  to  determine  its  condition,  and 
periodically,  thereafter,  to  determine  the  rate  of  ripening. 

Fast  ripening  may  be  accomplished  in  from  three  to  four  days  after  the 
green  fruit  is  placed  in  a  ripening  room  by  adjusting  the  temperatures  of 
the  room  until  the  pulp  temperature  reaches  about  70  F.  In  wanning  up 
cool  fruit,  quick  heating  is  recommended,  and  it  is  good  practice  to  use 
sufficient  heat  to  raise  the  average  fruit  temperature  at  the  rate  of  2  to 
3  deg  per  hour.  After  the  first  24  hours,  the  room  should  be  held  at  68  F 
until  the  fruit  is  colored  and  then  reduced  to  66  F  and  held  at  this  tem- 
perature. A  high  relative  humidity  of  from  90  to  95  per  cent  should  be 
maintained  until  the  bananas  show  color,  when  it  may  be  reduced  to  about 
80  per  cent.  High  humidity  is  important  during  the  warming  period. 
No  ventilation  should  be  used  until  the  fruit  has  colored,  after  which 
ventilation  at  a  rate  not  to  exceed  four  changes  per  hour  may  be  used  to 
assist  in  reducing  the  humidity  and  to  freshen  the  air  in  the  room.  If  the 
fruit  shows  slow  or  uneven  ripening  characteristics,  one  or  two  applica- 
tions of  ethylene  gas  of  approximately  1  cu  ft  per  1000  cu  ft  of  room  space 
may  be  used. 

Medium  speed  ripening  of  bananas  in  from  five  to  seven  days  may  be 
accomplished  by  holding  the  fruit  at  64  F.  The  humidity  and  ventilation 
control  should  be  the  same  as  for  fast  ripening.  A  treatment  with  ethy- 
lene gas  will  seldom  be  necessary.  For  slow  ripening  in  from  nine  to  ten 
days,  the  fruit  should  be  held  at  from  60  to  62  F.  Temperatures  below 
62  F  are  not  advisable  for  very  thin  fruit.  The  humidity  should  be  the 
same  as  for  fast  ripening,  and  ventilation  (up  to  3  or  4  air  changes  per 
hour)  should  be  used  provided  the  humidity  can  be  maintained.  Ethylene 
gas  treatment  will  not  be  required. 

For  holding  ripened  bananas,  temperatures  between  56  and  60  F  are 
recommended.  A  reduction  in  humidity  is  beneficial  in  toughening  the 
peel  and  reducing  the  mould,  but  too  low  a  humidity  will  cause  shrinkage. 
Although  exact  humidity  control  is  not  essential,  the  desirable  range  is 
between  75  and  80  per  cent. 

LUMBER  DRYING 

The  United  States  Forest  Products  Laboratory,  Madison,  Wis.,  has 

71 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


prepared  eleven  schedules2  for  the  kiln-drying  of  practically  all  kinds, 
types,  and  thicknesses  of  softwoods  and  hardwoods.  The  tables  given  in 
these  schedules  range  from  105  to  200  F  dry-bulb,  and  from  20  to  80  per 
cent  relative  humidity.  As  a  rule,  the  softer  the  wood,  the  higher  the 
average  temperature  used.  The  temperature  and  relative  humidity  in  a 
lumber  drying  kiln  are  varied  for  all  conditions,  starting  with  a  low  dry-- 
bulb and  a  high  relative  humidity  when  the  green  lumber,  containing  a 
large  percentage  of  moisture,  is  started  to  dry.  As  the  moisture  content 
of  the  lumber  decreases,  the  dry-bulb  temperature  of  the  kiln  is  increased, 
and  the  relative  humidity  reduced.  It  is  noted,  however,  that  perfect 
drying  does  not  necessarily  result  from  following  a  schedule,  and  that  an 
operator  must  be  trained  to  watch  the  condition  of  the  stock  in  the  kiln 
and  to  immediately  apply  a  remedy  if  he  sees  things  going  wrong. 

GREENHOUSES 

Table  3  lists  customary  dry-bulb  temperature  ranges  for  different 
types  of  plants  and  flowers  raised  in  greenhouses. 

TABLE  3.    CUSTOMARY  TEMPERATURES  FOR  DIFFERENT  TYPES  OF  GREENHOUSES 


TYPE  OF  HOUSE 

TEMPERATURE 
RANGE 
DEGFAHR 

TYPE  OP  HOUSE 

TEMPERATURE 
RANGE 
DEC  FA.HR 

Carnation              -                       

45  to  55 

Orchid,  cool  

50  to  55 

Conservatory  (general  collection) 

60  to  65 

Palm,  warm 

60  to  65 

Cool                                

45  to  50 

Palm,  cool   „    

50  to  55 

Cucumber  

65  to  70 

Propagating  

55  to  60 

Fern                   „    .              

60  to  65 

Rose-    „  

55  to  60 

Forcing 

60  to  65 

Sweet  pea 

45  to  50 

General  purpose 

55  to  60 

Tomato  ~        

65  to  70 

Lettuce 

40  to  45 

Tropical 

65  to  70 

Orchid,  warm 

65  to  70 

Violet  

40  to  45 

APPARATUS  FOR  INDUSTRIAL  CONDITIONING 

Apparatus  for  industrial  air  conditioning  may  be  divided  into  two 
distinct  groups,  namely,  (1)  humidifiers  for  increasing  the  moisture  con- 
tent of  the  air  and  for  producing  cooling  by  evaporation  and  (2)  dehu- 
midifiers  for  removing  moisture  from  the  air  and  for  producing  cooling  by 
contact  with  water  or  surfaces  at  a  lower  temperature  than  the  air. 

Strictly  speaking,  humidity  control  alone,  whether  it  involves  humidi- 
fication  or  dehumidification,  is  not  air  conditioning.  To  be  entitled  to  this 
classification  according  to  the  definition  in  Chapter  41,  the  process  should 
include  the  simultaneous  control  of  temperature,  humidity  and  air  motion. 

Industrial  humidifiers  may  be  divided  into  the  following  general 
types,  according  to  the  method  of  operation : 

1.  Direct,  which  spray  into  the  room. 

2.  Indirect,  which  introduce  moistened  air. 

3.  Combined  direct  and  indirect. 


-Technical  Note  Number  1 7o,  Forest  Products  Laboratory,  U.  S.  Forest  Service,  Madison,  Wis. 

72 


CHAPTER  3 — INDUSTRIAL  AIR  CONDITIONING 


Spray  Generation 

Spray  generation  is  obtained  by  (1)  atomization,  (2)  impact,  (3) 
hydraulic  separation,  and  (4)  mechanical  separation. 

Atomization  involves  the  use  of  a  compressed  air  jet  to  reduce  the  water 
particles  to  a  fine  spray.  With  the  impact  method,  a  jet  of  water  under 
pressure  impinges  directly  on  the  end  of  a  small  round  wire.  Where 
hydraulic  separation  is  employed,  a  jet  of  water  enters  a  cylindrical 
chamber  and  escapes  through  an  axial  port  with  a  rapid  rotation  which 
causes  it  immediately  to  separate  in  a  fine  cone-shaped  spray.  In  the 
mechanical  separation  process,  water  is  thrown  by  centrifugal  force  from 
the  surface  of  a  rapidly  revolving  disc  and  separates  into  particles  suf- 
ficiently small  to  be  utilized  in  certain  types  of  mechanical  humidifiers. 

Spray  Distribution 

Spray  distribution  is  obtained  by  (1)  air  jet,  (2)  induction,  and  (3)  fan 
propulsion. 

The  air  jet  which  generates  the  spray  in  atomizers  also  carries  the  spray 
through  a  space  sufficient  for  its  distribution  and  evaporation,  and  this 
method  of  distribution  is  termed  air  jet.  Where  distribution  is  obtained 
by  induction,  the  aspirating  effect  of  an  impact  or  centrifugal  spray  jet  is 
utilized  to  induce  a  current  of  air  to  flow  through  a  duct  or  casing,  and 
this  air  current  distributes  the  spray.  Fan  propulsion  obviously  consists 
of  the  utilization  of  fans  to  entrain  and  distribute  the  spray. 

Industrial  type  direct  humidifiers  are  commonly  classified  as  (1) 
atomizing,  (2)  high-duty,  (3)  spray  and  (4)  self-contained  or  centrifugal. 

Atomizing  Humidifiers 

There  are  several  types  of  atomizing  humidifiers,  all  of  which  rely  upon 
compressed  air  as  the  atomizing  and  distributing  agency,  similar  to  the 
familiar  method  used  in  ordinary  nasal  atomizers.  Compressed  air 
(ordinarily  about  30  Ib  per  square  inch)  is  supplied  from  a  centrally- 
located  air  compressor  through  pipe  lines  to  the  atomizing  units.  The  air 
lines  are  usually  horizontal  and  parallel  to  water  lines  which  supply 
water  by  gravity  from  a  float  tank.  The  water  in  the  tank  is  maintained 
at  a  constant  level  slightly  lower  than  the  outlets  of  the  atomizers  them- 
selves and  is  drawn  constantly  to  the  atomizer  by  aspiration  when  com- 
pressed air  is  supplied.  This  aspiration  ceases  and  the  flow  of  water  stops 
when  the  air  supply  is  cut  off.  The  water  should  not  be  supplied  under 
pressure  to  atomizers  because  of  the  possibility  of  leakage,  drip,  or  coarse 
spray  which  cannot  be  permitted  when  water  is  supplied  by  aspiration. 

High-Duty  Humidifiers 

Water  is  supplied  to  high-duty  humidifiers  under  high  pressure  (usually 
about  150  Ib  per  square  inch)  through  pipe  lines  from  a  centrally-located 
pumping  unit.  The  spray-generating  nozzle  which  is  of  the  impact  type 
is  located  in  a  cylindrical  casing,  A  drainage  pan  provides  for  the  collec- 
tion and  return  of  unevaporated  water  which  flaws  through  a  return  pipe 
to  a  filter  tank,  from  which  it  is  recirculated.  A  powerful  air  current  is 
forced  through  the  humidifier  by  means  of  a  fan  mounted  above  the  unit. 

73 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

The  air  enters  from  above,  is  drawn  through  the  head,  charged  with 
moisture,  and  cooled  to  the  wet-bulb  temperature.  It  then  escapes  from 
the  opening  below  at  a  high  velocity  in  a  complete  and  nearly  horizontal 
circle.  The  spray  is  quickly  evaporated  and  the  resulting  vapor  is  rapidly 
and  thoroughly  diffused.  This  effective  distribution  of  fine  spray  over 
the  maximum  possible  area  insures  complete  and  extremely  rapid  vapori- 
zation even  at  the  highest  humidities. 

Spray  Humidifiers 

This  type  of  humidifier  consists  of  an  impact  spray  nozzle  in  a  cylin- 
drical casing  with  a  drainage  pan  below  it.  The  aspirating  effect  of  the 
spray  nozzle  induces  a  moderate  air  current  through  the  casing  which 
distributes  the  entrained  spray.  The  general  method  of  circulating  and 
returning  the  water  is  similar  to  that  employed  for  high-duty  humidifiers. 
A  suitable  pump  and  centrally-located  filter  tank  are  required. 

The  spray  and  high-duty  types  of  humidifiers  have  many  features  in 
common  but  the  latter,  because  of  its  finer  spray  and  greater  capacity, 
is  often  considered  better  adapted  for  producing  high  humidities. 

Self-Contained  Humidifiers 

The  self-contained  or  centrifugal  humidifier  has  the  ability  to  generate 
and  distribute  spray  without  the  use  of  air  compressors,  pumps,  or  other 
auxiliaries.  These  may  be  used  either  singly  or  in  groups.  In  large 
installations,  where  suitable  connections  are  provided  to  permit  the 
cleaning  and  servicing  of  individual  units  without  affecting  the  room  as  a 
whole,  group  control  of  the  water  and  power  may  be  employed. 

Humidifiers  and  air  washers  are  also  described  in  Chapter  11. 

Where  large  quantities  of  power  are  generated  in  a  limited  space  and 
where  a  comparatively  high  relative  humidity  is  required,  it  is  often 
feasible  and  economical  to  use  a  combination  of  direct  and  indirect 
humidification.  The  indirect  humidification  provides  the  desired  quantity 
of  ventilation  and  cooling,  and  the  additional  direct  humidification  pro- 
vides for  increase  in  humidity  without  interfering  with  the  ventilation  or 
the  cooling  effected  by  the  indirect  system. 

In  general,  it  may  be  stated  that  direct  humidification  is  most  satis- 
factory where  high  humidities  are  desired  but  where  little  cooling,  ven- 
tilation or  air  motion  is  required.  Therefore,  the  indirect  system  is  most 
applicable  where  either  low  or  high  relative  humidities  are  desired  with 
maximum  cooling  and  ventilation  effect.  For  conditions  that  require  an 
unusually  large  amount  of  heat  to  be  absorbed  by  ventilation,  together 
with  the  maintenance  of  high  humidities,  it  is  often  preferable  to  make 
use  of  the  combination  system  of  indirect  and  direct  humidification.  If 
the  indirect  system  alone  were  used  it  would  mean  an  unusually  large 
volume  of  air  to  be  handled,  which  might  interfere,  due  to  air  motion, 
with  production,  even  though  it  would  result  in  greater  cooling  effect.  If 
direct  humidification  alone  were  used,  no  ventilation  would  be  obtained, 
with  consequently  higher  room  temperatures. 

Dehumidifiers,  which  are  similar  in  design  and  appearance  to  indirect 
humidifiers  and  air  washers,  are  described  in  Chapter  11.  The  main 
differences  are  found  in  the  internal  construction  of  the  dehumidifier,  in 

74 


CHAPTER  3 — INDUSTRIAL  AIR  CONDITIONING 


the  use  of  refrigeration  or  of  heat  as  required  for  controlling  the  water 
temperature,  and  in  differences  in  the  general  methods  of  control. 


PROBLEMS  IX  PRACTICE 

1  •  A  condition  of  75  F  dry -bulb  temperature  and  55  per  cent  relative  humidity 
is  being  maintained  in  a  cigarette  manufacturing  department.    What  will  be 
the  regain  and  moisture  content  of  the  tobacco? 

The  regain,  from  Table  1  =  17,75  per  cent. 

~,          .  17.75  X  100 

The  moisture  content  =  r^    .    .,-,-;-  =  lo.l  per  cent. 
100  4-  17./O 

2  •  A  1-lb  sample  taken  from  a  100-lb  batch  of  material  is  found  to  have  a  bone 
dry  weight  of  0.89  Ib.     This  material  is  to  be  processed  under  atmospheric 
conditions  which  should  produce  a  regain  of  15  per  cent.    Compute  the  finished 
weight  for  each  original  100-lb  batch. 

Let  W  equal  the  number  of  pounds  of  moisture  in  a  finished  batch. 

W  „  ,15 

gg- regain  -lo  per  cent  -jgg 

W  =  13.35 

89  +  13.35  =  102.35  Ib  finished  weight. 

3  •  A  bundle  of  sea  island  cotton  is  found  to  have  a  bone  dry  weight  of  9.26  Ib- 
What  is  the  proper  relative  humidity  at  75  F  to  produce  a  weight  of  10  Ib  at 
equilibrium? 

Desired  conditioned  weight  =  10.00  Ib 
Bone  dry  weight  =    9.26  Ib 

Weight  of  moisture  required  =    0.74  Ib 

074 
Regain  =  -~  X  100  =  7.9  per  cent. 

From  Table  1,  the  proper  relative  humidity  required  is  60  per  cent. 

4  •  Compute  tlie  bone  dry  weight  of  1000  Ib  of  manila  rope  which  has  been, 
stored  for  a  considerable  period  of  time  in  a  conditioned  room  at  75  F  dry-bulb 
temperature  and  50  per  cent  relative  humidity. 

Assuming  that  this  material  has  come  to  equilibrium  under  the  atmospheric  conditions 

given,  Table  1  shows  a  regain  of  8.5  per  cent. 

Let  W  equal  the  total  weight  of  moisture  in  pounds. 

1000  —  W  —  bone  dry  weight  in  pounds. 

=  regain  =8.5  per  cent 


1000  -  W         ^  ^  100 

W  =  78.3  Ib  moisture 
1000  -  78.3  =  921.7  Ib  bone  dry  weight. 

5  •  An  egg  evaporating  plant  wishes  to  dry  2000  Ib  of  egg  whites  (85  per  cent 
water)  to  crystalline  form  each  24  hours*  The  nmyimmm  permissible  air  de- 
livery temperature  in  the  dryer  is  140  F.  What  air  volume  will  be  required, 
assuming  that  outside  air  is  at  95  F  dry-bulh  and  78  F  wet-bulb  and  that  air 
leaves  the  dryer  70  per  cent  saturated? 

Moisture  to  be  removed  =  2000  X  0.85  =  1700  Ib.  Using  psychroroetric  chart  and 
starting  at  the  intersectioH  of  the  vertical  95  F  dry-bulb  temperature  line  and  the  45  per 
cent  humidity  Ene,  move  horizontally  to  tlie  right  to  the  intersection  with  the  140  F 
vertical  temperature  line  at  10  per  cewt  relative  haxmdHy ;  then  inove  along  the  constant 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

heat  'or  wet-bulb  line;  to  its  intersection  with  the  70  per  cent  relative  humidity  curve 
and  read  94  F  dry-bulb,  which  will  be  the  temperature  of  the  air  leaving  the  dryer. 

Moisture  per  cubic  foot  at  94  F  and  70  per  cent  relative  humidity  =  11.8  grains 
Moisture  per  cubic  foot  at  95  F  and  78  F  wet-bulb  =    8.0  grains 

Moisture  added  per  cubic  foot  of  air  handled  =     3.8  grains 

1700  X  7000 


No  allowance  is  made  for  heat  lost  in  the  transmission  to  and  from  the  dryer  or  for  the 
heat  required  to  raise  the  product  from  its  entering  temperature  to  that  maintained  in  the 
dryer.  This  would  necessitate  a  trial  and  error  solution  common  to  all  drying  problems. 

6  •  It  is  proposed  to  install  a  central  fan  type  air  conditioning  system  com- 
prised of  fan,  air  washer,  filters,  and  heating  coils  to  provide  ventilation  and  to 
maintain  proper  humidity  in  a  small  library  during  periods  of  winter  operation. 
The  heat  loss  has  been  estimated  at  450,000  Btu  per  hour  in  maintaining  a 
condition  of  72  F  dry-bulb  and  45  per  cent  relative  humidity.    Assuming  that 
the  air  washer  completely  saturates  the  air,  what  must  be  the  leaving  dry-and 
wet-bulb  temperatures  to  provide  the  required  condition? 

49.85  F  is  the  dew-point  temperature  corresponding  to  the  stated  required  condition, 

7  •  Assuming  a  maximum  permissible  air  delivery  temperature  of  100  F  in 
Question  6,  what  air  volume  will  be  required? 

450,000  X  55.2 


(100  -  72)  X  60 


14,800  cfm. 


8  •  If  in  Questions  6  and  7  it  is  assumed  that  winter  humidity  control  will 
consist  simply  of  a  dew-point  thermostat  at  the  exit  of  the  air  washer,  control- 
ling the  dew-point  temperature  by  operating  automatic  dampers,  and  thereby 
proportioning  the  respective  volumes  of  outside  and  recirculated  air  admitted: 

a.  What  volume  of  air  should  be  recirculated? 

b.  What  volume  of  air  will  be  exfiltrated  from  the  buildings? 

c.  What  reheating  capacity  will  be  required? 

a.  Btu  per  pound  at  72  F  and  '45  per  cent  relative  humidity  =  25.38 
Btu  per  pound  at  0  F  (assumed  saturated)  =    0.85 
Btu  per  pound  at  49.85  F  saturated  =  20.11 

Recirculated  air  =  ^5  38  ~  0  85)  X  14'8°°  =  11'6°°  cfm* 

b.  The  same  volume  as  is  introduced  as  fresh  outside  air,  namely, 

14,800  -  11,600  =  3200  cfm. 

c.  The  reheaters  must  be  of  such  capacity  as  to  reheat  the  volume  of  air 
handled  from  49.85  (the  dew-point)  to  100  F. 

14,800  X  (100  -  49.85)  X  60 


55.2 


=  808,000  Btu  per  hour. 


76 


Chapter  4 

NATURAL  VENTILATION 

Wind  Forces,  Stack  Effect,  Openings,  Windows,  Doors,  Skylights, 

Roof  Ventilators,  Stacks,  Principles  of  Control,   General  Rules, 

Measurements,  Dairy  Barn   Ventilation,    Garage   Ventilation 

VENTILATION  by  natural  forces,  supplemented  in  certain  cases 
with   mechanical   forces,    finds   extensive   application   in   industrial 
plants,  public  buildings,  schools,  dwellings,  garages,  and  in  farm  buildings. 
The  natural  forces  available  for  the  displacement  of  air  in  buildings  are 
the  wind  and  the  difference  in  temperature  of  the  air  inside  and  outside 
the  building.     The  arrangement  and  control  of  ventilating  openings 
should  be  such  that  the  two  forces  act  cooperatively  and  not  in  opposition, 

Wind  Forces 

In  considering  the  use  of  natural  wind  forces  for  the  operation  of  a 
ventilating  system,  account  must  be  taken  of  (1)  average  and  minimum 
wind  velocities,  (2)  wind  direction,  (3)  seasonal,  daily  and  hourly  varia- 
tions in  wind  velocity  and  direction,  and  (4)  local  wind  interference  by 
buildings  and  trees. 

Table  1,  Chapter  8,  gives  values  for  the  average  summer  wind  velocities 
and  the  prevailing  wind  directions  in  various  localities  throughout  the 
United  States,  while  Table  2,  Chapter  7,  lists  similar  values  for  the  winter. 
In  almost  all  localities  the  summer  wind  velocities  are  lower  than  those  in 
the  winter,  and  in  about  two-thirds  of  the  localities  the  prevailing  direc- 
tion is  different  during  the  summer  and  winter.  While  average  wind 
velocities  are  seldom  below  5  mph,  there  are  many  hours  in  each  month 
during  which  the  wind  velocity  is  from  3  to  5  mph,  even  in  localities  where 
the  seasonal  average  is  considerably  above  5  mph.  There  are  relatively 
few  places  where  the  hourly  wind  velocity  falls  much  below  3  mph  for 
more  than  10  daylight  hours  per  month.  Usually  a  natural  ventilating 
system  should  be  designed  to  operate  satisfactorily  with  a  wind  velocity 
of  3  to  6  mph,  depending  on  locality. 

The  following  formula  may  be  used  for  calculating  the  quantity  of  air 
forced  through  ventilation  openings  by  the  wind,  or  for  determining  the 
proper  size  of  such  openings: 

Q  =  EA  V  (1) 

where 

Q  =  air  flow  in  cubic  feet  per  minute.  _ 

A  —  free  area  of  inlet  (or  outlet)  openings  in  square  feet. 

V  —  wind  velocity  in  feet  per  minute, 

—  miles  per  hour  X  88. 
E  =  effectiveness  of  openings. 

(R  sfconld  be  taken  at  from  50  to  60  per  cent  if  the  inlet  openings  face  the  wind  and  from  25  to  35  per 
cent  if  the  infet  openinigs  receive  tfoe  wirad  at  an  angle.) 

in 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

If  outlet  openings,  where  air  leaves  a  building,  are  smaller  than  inlet 
openings,  where  air  enters  a  building,  the  air  will  be  less  effective  than 
indicated  by  the  constant  E. 

The  accuracy  of  the  results  obtained  by  the  use  of  Formula  1  depends 
upon  the  placing  of  the  openings,  as  the  formula  assumes  that  ventilating 
openings  have  a  flow  coefficient  slightly  greater  than  that  of  a  square-edge 
orifice.  If  the  openings  are  not  advantageously  placed  with  respect  to  the 
wind,  the  flow  per  unit  area  of  the  openings  will  be  less,  and  if  unusually 
well  placed,  the  flow  will  be  slightly  more  than  that  given  by  the  formula. 
Inlets  should  be  placed  to  face  directly  into  the  prevailing  wind,  while 
outlets  should  be  placed  in  one  of  the  following  four  places : 

1.  On  the  side  of  the  building  directly  opposite  the  direction  of  the  prevailing  wind. 

2.  On  the  roof  in  the  low  pressure  area  caused  by  the  jump  of  the  wind  (see  Fig.  1). 

3.  In  a  monitor  on  the  side  opposite  from  the  wind. 

4.  In  roof  ventilators  or  stacks  exposed  to  the  full  force  of  the  wind1. 

Forces  due  to  Stack  Effect2 

The  stack  effect  produced  within  a  building  is  due  to  the  difference  in 
weight  of  the  warm  column  of  air  within  the  building  and  the  cooler  air 
outside.  The  flow  due  to  stack  effect  is  proportional  to  the  square  root 
of  the  draft  head,  or  approximately: 


Q  -  9.4  A  V  H  (ti  -  *2)  (2) 

where 

Q  —  air  flow  in  cubic  feet  per  minute. 

A  =  free  area  of  inlets  or  outlets  (assumed  equal)  in  square  feet. 
H  —  height  from  inlets  to  outlets,  in  feet. 

ti  —  average  temperature  of  indoor  air  in  height  H,  in  degrees  Fahrenheit. 
/2  =  temperature  of  outdoor  air,  in  degrees  Fahrenheit. 

9.4  ss  constant  of  proportionality,  including  a  value  of  65  per  cent  for  effectiveness  of 
openings.  This  should  be  reduced  to  50  per  cent  (constant  =  7.2)  if  conditions 
are  not  favorable. 

The  height  between  inlets  and  outlets  should  be  the  maximum  which 
the  building  construction  will  allow. 

In  some  cases  the  necessary  air  flow  will  be  known  from  the  require- 
ments of  the  building  occupancy,  and  the  area  necessary  for  certain 
assumed  temperature  differences  may  be  calculated.  Or  the  areas  may 
be  fixed  by  the  building  construction,  and  the  maximum  air  flow  for 
various  differences  between  indoor  and  outdoor  temperatures  may  be 
calculated.  In  any  case,  the  conditions  which  give  the  minimum  air  flow 
are  those  which  control  the  design,  as  the  system  must  have  ample 
capacity  even  under  the  most  unfavorable  conditions  which  are  those  of 
mild  or  warm  weather. 

TYPES  OF  OPENINGS 

The  engineering  problems  of  a  natural  ventilation  system  consist  of  the 
design,  location,  and  control  of  ventilating  openings  to  best  utilize  the 


'See  Airation  of  Industrial  Buildings,  by  W.  C.  Randall  (A.S.H.V.E.  TRANSACTIONS,  Vol.  34,  1928). 

2See  Neutral  Zone  in  Ventilation,  by  J.  E.  Emswiler  (A.S.H.V.E.  TRANSACTIONS,  Vol.  32,  1926),  and 
Predetermining  Airation  of  Industrial  Buildings,  by  W.  C.  Randall  and  E.  W.  Conover  (A.S.H.V.E.  TRANS- 
ACTIONS, Vol.  37,  1931). 

78 


CHAPTER  4 — NATURAL  VENTILATION 


natural  ventilation  forces,  in  accordance  with  the  requirements  of  build- 
ing occupancy.    The  types  of  openings  may  be  classified  as: 

1.  Windows,  doors,  monitor  openings,  and  skylights. 

2.  Roof  ventilators. 

3.  Stacks  connecting  to  registers. 

4.  Specially  designed  inlet  or  outlet  openings. 

Windows,  Doors  and  Skylights 

Windows  have  the  advantage  of  transmitting  light,  as  well  as  providing 
ventilating  area  when  open.    Their  movable  parts  are  arranged  to  open  in 


FIG.  1.    THE  JUMP  OF  WIND  FROM  WINDWARD  FACE  OF  BUILDING.    (A— LENGTH  or 

SUCTION  AREA;  B — POINT  OF  MAXIMUM  INTENSITY  OF  SUCTION; 

C — POINT  OF  MAXIMUM  PRESSURE) 

various  ways;  they  may  open  by  sliding  as  in  the  ordinary  double-hung 
windows,  by  tilting  on  horizontal  pivots  at  or  near  the  center,  or  by 
swinging  on  pivots  at  the  top  or  bottom.  Whatever  the  form  and  type  of 
window  used,  the  amount  of  dear  area  that  can  be  made  available  is  the 
factor  of  greatest  importance  in  ventilation. 

All  types  of  sash  (double-hung,  top,  center  or  bottom  horizontal  pivoted, 
or  vertical  pivoted)  have  about  the  same  air  flow  capacity  for  the  same 
clear  area.  Air  leakage  through  dosed  windows  is  important  during  high 
winds  (Chapter  6). 

7§ 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

The  proper  distribution  of  air  in  occupied  spaces  is  an  element  almost 
as  important  as  that  of  sufficient  air  quantity.  Advantageous  pivoting  of 
sash  is  very  useful  for  securing  good  air  distribution.  Deflectors  are  some- 
times used  for  the  same  purpose,  and  these  devices  should  be  considered  a 
part  of  the  ventilation  system. 

Door  openings  are  seldom  included  in  the  ventilation  calculations, 
though  they  may  be  of  great  value  for  extreme  summer  conditions,  and 
should  be  considered  in  this  connection  as  well  as  in  garage  design. 

Skylight  and  monitor  openings  are  of  importance  as  these  and  the  roof 
ventilators  are  outlets,  while  the  lower  windows  are  usually  inlets  on  the 
windward  side  and  outlets  on  the  leeward  side.  In  general  the  areas  of 
inlets  and  of  outlets  should  be  about  equal.  It  is  important  to  make  a 
check  on  this  ratio  in  any  installation,  as  any  great  excess  of  area  of  one 
set  of  openings  over  another  means  waste  opening  area.  The  operating 
devices  used  for  sash,  monitors,  skylights  and  roof  ventilators  should  be 
well  selected  as  poor  operating  devices  may  defeat  the  entire  design. 

Roof  Ventilators 

The  function  of  a  roof  ventilator  is  to  provide  a  storm  and  weather 
proof  air  outlet,  which  is  sensitive  to  wind  action  for  producing  additional 
flow  capacity,  and  at  the  same  time  is  subject  to  manual  or  automatic 
control  by  suitable  dampers.  The  capacity  of  a  ventilator  at  a  constant 
wind  velocity  and  temperature  difference,  depends  upon  four  things: 
(1)  its  location  on  the  roof,  (2)  the  resistance  it  offers  to  air  flow,  (3)  the 
area  and  location  of  openings  provided  for  air  inflow  at  a  lower  level,  and 
(4)  the  ability  of  the  ventilator  head  to  utilize  the  kinetic  energy  of  the 
wind  for  inducing  flow  by  centrifugal  or  ejector  action.  Frequently  one 
or  more  of  these  capacity  factors  is  overlooked  in  a  ventilator  installation. 

For  maximum  flow  induction,  a  ventilator  should  be  located  on  that 
part  of  the  roof  which  receives  the  full  wind  without  interference.  (See 
Fig.  1.)  This  does  not  mean  that  no  ventilators  are  to  be  installed  within 
the  suction  region  created  by  the  wind  jumping  over  the  building,  or  in  a 
light  court,  or  on  a  low  building  between  two  high  buildings.  Ventilators 
are  highly  effective  in  such  low-pressure  areas,  but  their  ejector  action, 
caused  by  wind  velocity,  is  of  little  importance  in  these  locations,  and 
hence  their  size  should  be  increased  proportionally. 

Ventilator  resistance  depends  on  (1)  type  of  inlet,  (2)  area  of  openings 
and  passages,  and  (3)  number  of  turns  or  changes  of  direction  of  the  air 
flow.  The  inlet  grille,  if  any,  should  have  ample  free  area,  and  the  venti- 
lator should  always  be  provided  with  a  taper-cone  inlet  in  order  to  produce 
the  effect  of  a  bell-mouth  nozzle  (flow  coefficient  0.97)  rather  than  that  of 
a  square-entrance  orifice  (flow  coefficient  0.60) .  In  other  words,  the  grilles 
should  be  oversize  as  compared  with  the  ventilator,  and  they  should  be 
connected  by  tapering  collars.  If  the  ventilator  head  construction 
produces  changes  in  the  direction  of  air  flow,  the  area  of  the  flow  passages 
should  be  increased  accordingly. 

Air  inlet  openings  at  lower  levels  in  the  building  are  of  course  necessary 
for  the  economical  use  of  ventilator  capacity.  The  inlet  openings  should 
be  at  least  equal  to,  and  preferably  twice  as  great  as  the  combined  throat 
areas  of  all  roof  ventilators.  The  air  discharged  by  a  roof  ventilator 

80 


CHAPTER  4 — NATURAL  VENTILATION 


depends  on  wind  velocity  and  temperature  difference,  but  due  to  the  four 
capacity  factors  already  mentioned,  no  simple  formula  can  be  devised  for 
expressing  ventilator  capacity. 

Several  types  of  roof  ventilators  are  shown  in  Figs.  2  to  11.  These  may 
be  classified  as  stationary,  Figs.  2  to  6,  pivoted  or  oscillating,  Figs.  7  to  9, 
or  rotating,  Figs.  10  and  11.  When  selecting  unit  ventilators,  some 
attention  should  be  paid  to  ruggedness  of  construction,  storm-proofing 
features,  dampers  and  damper  operating  mechanisms,  possibilities  of 
noise  from  dampers  or  other  moving  parts,  and  possible  maintenance 
costs. 

It  should  be  kept  in  mind  that  a  suitable  combination  of  roof  venti- 
lators with  mechanical  ventilation  frequently  offers  the  best  solution  of  a 
ventilating  problem.  The  natural  ventilation  units  may  be  used  to  sup- 
plement power  driven  supply  fans,  and  under  favorable  weather  con- 
ditions it  may  be  possible  to  shut  down  the  power  driven  units.  Where 
low  operating  costs  are  very  important,  such  a  combination  has  great 
advantages.  Roof  ventilators  with  built-in  electric  fans  are  attracting 
increased  attention  because  they  combine  the  advantages  of  low  instal- 
lation and  operating  cost  with  those  of  continuous  service. 

Controls 

In  connection  with  any  combination  between  natural  and  fan  venti- 
lation, the  controls  are  of  importance.  Both  the  fans  and  the  ventilator 
dampers  may  be  controlled  by  some  combination  of  three  methods: 
(1)  hand  operation,  (2)  thermostat  operation,  and  (3)  control  by  wind 
velocity.  The  thermostat  station  may  be  located  anywhere  in  the 
building,  or  it  may  be  located  within  the  ventilator  itself.  The  purpose  of 
wind  velocity  control  is  to  obtain  a  definite  volume  of  exhaust  regardless 
of  the  natural  forces,  the  fan  motor  being  energized  when  the  natural 
exhaust  capacity  falls  below  a  certain  minimum,  and  again  shut  off  when 
the  wind  velocity  rises  to  the  point  where  this  minimum  volume  can  be 
supplied  by  natural  forces. 

Stacks 

Stacks  are  really  chimneys  and  utilize  both  the  inductive  effect  of  the 
wind  and  the  force  of  temperature  difference  (the  so-called  gravity  action). 
While  their  openings  projecting  above  the  roof  are  not  provided  with  any 
special  construction  for  developing  suction  by  the  action  of  the  wind,  the 
plain  vertical  opening  is  also  effective  in  this  respect.  Like  the  roof 
ventilator,  the  stack  outlet  should  be  located  so  that  the  wind  may  act 
upon  it  from  any  direction. 

Stacks  are  applicable  particularly  in  the  case  of  schools,  apartments, 
residences  and  small  office  buildings.  Partitions  interfere  with  general 
air  circulation,  and  some  type  of  outlet  from  each  room  is  necessary.  If 
the  building  is  not  too  tall,  and  the  requirements  of  occupancy  are  moder- 
ate, a  system  of  stacks  with  registers  in  each  room  may  be  more  eco- 
nomical than  a  system  of  mechanical  ventilation  employing  fans.  In 
making  the  comparison,  however,  the  building  space  occupied  by  the 
stacks  should  be  considered. 

With  little  or  no  wind,  chimaey  effept  or  temperature  difference  will 
produce  outflow  through  the  stacks  and  an  equal  inflow  through  windows 

81 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


(     1 

I 

/ 

^*  ' 

A 

— 

^™ 

71 

JP 

f 

\*      ^ 

N 

r      V_ 

af 

£.— 

_J>    ^ 

X 

FIG.  2 


FIG.   3 


FIG.  4 


FIG.  5  FIG.  6 

Six.  COMMON  TYPES  OF  STATIONARY  VENTILATORS 


FIG.  7  FIG.  8  FIG.  9 

THREE  TYPICAL  OSCILLATING  VENTILATORS 

82 


CHAPTER  4 — NATURAL  VENTILATION 


in  all  sides  of  the  building.  With  wind,  the  inductive* force  at  the  top  of 
ventilating  shafts  is  more  powerful  than  that  on  the  leeward  side  of  the 
building,  so  that  air  is  drawn  in  through  leeward  openings  by  a  combina- 
tion of  the  forces  of  wind  and  temperature  difference.  On  the  windward 
side,  the  direct  forcing  pressure  of  the  wind  is  of  course  added  to  the 
temperature  difference  effect.  Thus  forces  are  available  for  causing  in- 
flow at  practically  every  window  of  such  a  building.  Adequacy  of  stack 
size  must,  of  course,  be  provided. 

PRINCIPLES  OF  AIR  FLOW  CONTROL 

The  air  flow  through  a  ventilation  opening  depends  on  the  two  factors 
already  discussed,  namely,  (1)  the  natural  forces  available,  (2)  the  open- 
ings available,  and  the  resistance  to  flow  offered  by  these  openings.  The 
design  problem  includes,  of  course,  a  determination  of  the  desired  air 


/  Propelling  blai 


FIG.  10. 


SE.OTIOM 


ROTATING  VENTILATORS 


FIG.  11. 


quantity  and  distribution  in  order  that  the  openings  may  be  properly 
placed. 

The  purpose  of  ventilation  is  to  carry  off  either  excess  heat  or  air 
impurities,  and  the  desired  air  quantities  depend  upon  the  amount  of  heat 
or  of  impurities  present.  The  amount  of  heat  can  be  determined,  in  the 
case  of  forge  shops  for  example,  from  the  amount  of  fuel  burned,  which  in 
turn  is  based  upon  the  production  capacity  for  which  the  building  is 
being  designed.  In  the  case  of  foundries,  the  heat  given  off  by  the  metal 
in  cooling  from  the  molten  state  can  be  used.  In  some  instances,  not  all 
of  the  heat  may  be  dissipated  to  the  air,  but  a  fair  estimate  of  the  amount 
to  be  removed  by  the  air  can  usually  be  made. 

The  next  step  is  to  select  the  temperature  difference  to  be  maintained. 
Knowing  the  amount  of  heat  to  be  removed  and  having  selected  a 
desirable  temperature  difference,  the  amount  of  air  to  be  passed  through 
the  building  per  minute  to  maintain  this  temperature  difference  can  be 
determined  by  means  of  the  following  equation : 


H 


where 


cQD 

V 


(3) 


c  ~  0.24  =  specific  heat  of  air. 

V  —  specific  volume  of  the  air,  cubic  feet  per  pound,  about  13.5.  (See  Chapter  41.) 

83 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

H  =  heat  to  be  carried  off,  in  Btu  per  minute. 

Q  —  air  flow  in  cubic  feet  per  minute. 

D  =  inlet-outlet  temperature  difference  in  degrees  Fahrenheit. 

For  disposing  of  air  impurities,  the  required  air  flow  must  be  such  that 
the  outside  air  will  dilute  the  impurities  to  a  degree  that  they  are  no 
longer  objectionable.  For  human  occupancy,  such  as  in  auditoriums  and 
classrooms,  10  cfm  per  person  is  usually  taken  as  the  minimum  of  outside 
air  necessary  for  ventilation  (see  Chapter  2).  For  garage  ventilation, 
sufficient  air  must  be  admitted  to  dilute  the  carbon  monoxide  content  of 
the  indoor  air  to  1  in  10,000  (see  Garage  Ventilation  in  this  Chapter). 

Air  quantity  and  quality  are  not  the  only  requirements.  For  human 
occupancy,  air  distribution  is  important.  In  ventilation  the  air  distribu- 
tion is  almost  entirely  a  matter  of  the  number,  the  design,  and  the  location 
of  inlets  and  outlets.  In  locating  openings,  special  precautions  should  be 
taken  against  the  formation  of  dead  air  spaces  or  pockets  within  the  zone 
of  occupancy. 

Suggested  methods  for  estimating  the  air  flow  due  to  temperature 
difference  alone  and  to  wind  alone  have  already  been  given.  It  must  be 
remembered  that  when  both  forces  are  acting  together,  even  without 
interference,  the  resulting  air  flow  is  not  equal  to  the  sum  of  the  two 
estimated  quantities.  The  same  openings  have  been  assumed  in  both 
cases,  and  since  the  resistance  to  flow  through  the  openings  varies  ap- 
proximately with  the  square  of  the  velocity3,  this  resistance  becomes  a 
limiting  factor  as  the  flow  through  the  openings  is  increased. 

Recent  investigations1* 2  show  that  the  total  flow  is  only  10  per  cent 
above  the  flow  caused  by  the  greater  force  when  the  two  forces  are  nearly 
equal,  and  this  percentage  decreases  rapidly  as  one  force  increases  above 
the  other.  Tests  on  roof  ventilators  indicate  that  this  is  too  conservative 
in  the  direction  of  low  total  flow  quantities,  but  there  is  in  any  case  a 
large  judgment  factor  involved.  The  wind  velocity  and  direction,  the 
outdoor  temperature,  or  the  indoor  activities  cannot  be  predicted  with 
certainty,  and  great  refinement  in  calculations  is  therefore  not  justified. 
When  designing  for  winter  conditions,  an  added  variable  is  the  heat  lost 
by  direct  flow  through  walls  and  windows  and  by  infiltration. 

Example  1.  Assume  a  drop  forge  shop,  200  ft  long,  100  ft  wide,  and  30  ft  high.  The 
cubical  content  is  600,000  cu  ft,  and  the  height  of  the  air  outlet  over  that  of  the  inlet  is 
30  ft.  Oil  fuel  of  18,000  Btu  per  Ib  is  used  in  this  shop  at  the  rate  of  15  gal  per  hour 
(7.75  Ib  per  gal) .  Temperature  differences  are  10  F  in  summer  and  30  F  in  winter,  and 
the  wind  velocity  is  5  mph  in  summer  and  8  mph  in  winter.  What  is  the  necessary  area 
for  the  inlets  and  outlets,  and  what  is  the  rate  of  air  flow  through  the  building? 

Solution.  The  system  must  be  designed  for  the  summer  conditions  as  these  are  the 
more  severe.  The  heat  to  be  removed  per  minute  is: 

H  -  ^-  X  7.75  X  18,000  -  34,875  Btu. 

uu 

By  Equation  3,  the  air  flow  required  to  remove  this  heat  with  a  temperature  difference 
of  10  deg  is: 

VH        13.5  X  34,875       . 
Q  =  -& 0.24X10       =  1 


This  is  true  for  turbulent  flow  only.  It  would  be  more  correct  to  state  that  the  resistance  varies  approxi- 
mately with  V2  for  high  to  moderate  velocities,  with  F1'8  for  moderate  to  low  velocities,  and  with  the  first 
power  of  the  velocity  for  very  low  velocities  through  small  openings. 

84 


CHAPTER  4 — NATURAL  VENTILATION 


This  is  equal  to  19.6  air  changes  per  hour.  The  assumption  is  made  that  the  average 
temperature  difference  between  indoors  and  outdoors  is  the  same  as  the  temperature  rise 
of  the  air  from  the  inlet  opening  to  the  outlet  opening.  Actually,  the  latter  difference  is 
larger  and  so  the  value  of  19.6  air  changes  per  hour  is  conservative  as  it  allows  for  more 
cooling  than  is  necessary  for  an  average  temperature  difference  of  10  deg. 

If  196,172  cfm  are  to  be  circulated  by  the  force  of  the  temperature  difference  alone,  the 
area  of  opening  would  be,  by  Equation  2: 

196,172 


If  this  area  of  openings  were  provided,  a  wind  velocity  of  5  mph,  acting  alone,  would 
produce  a  flow  according  to  Equation  1,  of: 

<2  «  EA  V  =  0.50  X  1,205  X  5  X  88  =  265,100  cfm. 

If  the  inlet  openings^do  not  face  the  wind,  but  are  at  an  angle  with  it,  about  half  this 
amount  may  be  considered  to  flow. 

A  factor  of  judgment  must  now  be  exercised  in  making  the  selection  of 
the  area  of  openings  to  be  specified.  Apparently  1205  sq  ft  are  a  very 
generous  allowance  because  either  a  direct  wind  of  5  mph  or  an  average 
temperature  difference  of  10  deg  acting  alone  will  more  than  suffice  to 
carry  away  the  heat,  and  when  the  two  forces  are  acting  together,  the 
system  may  have  an  excess  capacity  of  25  per  cent  to  50  per  cent,  especially 
if  the  outlets  are  made  up  partially  of  roof  ventilators  which  employ  the 
force  of  the  wind  for  producing  a  suction  effect.  On  the  other  hand,  the 
wind  may  at  times  come  from  an  unfavorable  direction,  or  its  velocity 
may  fall  below  5  mph  or  the  building  construction  may  not  permit  a  full 
2400  sq  ft  of  inlet  window  area  and  an  equal  amount  of  monitor  or  roof 
ventilator  outlet  area.  In  case  the  two  sets  of  openings  are  not  equal, 
their  effectiveness  is  reduced. 

From  this  example  it  must  be  apparent  that  while  formulas  may 
furnish  a  reliable  guide,  the  final  solution  of  a  problem  of  natural  venti- 
lation requires  a  common  sense  analysis  of  local  conditions  to  supplement 
and  to  modify  the  dictates  of  the  formulas. 

GENERAL  RULES 

A  few  of  the  important  requirements  in  addition  to  those  already 
outlined  are: 

1.  Inlet  openings  should  be  well  distributed,  and  should  be  located  on  the  windward 
side  near  the  bottom,  while  outlet  openings  are  located  on  the  leeward  side  near  the  top. 
Outside  air  will  then  be  supplied  to  the  zone  of  occupancy. 

2.  Direct  short  circuits  between  openings  on  two  sides  at  a  high  level  may  clear  the 
air  at  that  level  without  producing  any  appreciable  ventilation  at  the  level  of  occupancy. 

3.  Roof  ventilators  should  be  located  20  to  40  ft  apart  each  way  and  preferably  on 
the  ridge  of  the  roof.    The  closer  spacings  are  used  when  ventilating  rooms  with  low 
ceilings. 

4.  Greatest  flow  per  square  foot  of  total  opening  is  obtained  by  using  inlet  and  outlet 
openings  of  nearly  equal  areas. 

5.  In  an  industrial  building  where  furnaces,  that  give  off  heat  and  fumes,  are  to  be 
installed,  it  is  better  to  locate  them  in  the  end  of  the  building  exposed  to  the  prevailing 
wind.  The  strong  suction  effect  of  the  wind  at  the  roof  aear  the  windwajrd  end  will  then 
cooperate  with  temperature  difference,  to  provide  for  the  most  active  and  satisfactory 
removal  of  the  heat  and  gas  laden  air. 

85 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

6.  In  case  it  is  impossible  to  locate  furnaces  in  the  windward  end,  that  part  of  the 
building  in  which  they  are  to  be  located  should  be  built  higher  than  the  rest,  so  that 
the  wind,   in  splashing  therefrom  will  create  a  suction.  The  additional  height  also 
increases  the  effect  of  temperature  difference  to  cooperate  with  the  wind. 

7.  In  the  use  of  monitors,  windows  on  the  windward  side  should  usually  be  kept 
closed,  since,  if  they  are  open,  the  inflow  tendency  of  the  wind  counteracts  the  outflow 
tendency  of  temperature  difference.   Openings  on  the  leeward  side  of  the  monitor  result 
in  cooperation  of  wind  and  temperature  difference. 

8.  In  order  that  the  force  of  temperature  difference  may  operate  to  maximum  advan- 
tage, the  vertical  distance  between  inlet  and  outlet  openings  should  be  as  great  as 
possible.    Openings  in  the  vicinity  of  the  neutral  zone  are  less  effective  for  ventilation. 

9.  In  order  that  temperature  difference  may  produce  a  motive  force,  there  must  be 
vertical  distance  between  openings.  That  is,  if  there  are  a  number  of  openings  available 
in  a  building,  but  all  are  at  the  same  level,  there  will  be  no  motive  head  produced  by 
temperature  difference,  no  matter  how  great  that  difference  might  be. 

10.  In  the  design  of  window  ventilated  buildings,  where  the  direction  of  the  wind  is 
quite  constant  and  dependable,  the  orientation  of  the  building  together  with  amount 
and  grouping  of  ventilation  openings  can  be  readily  arranged  to  take  full  advantage  of 
the  force  of  the  wind.    On  the  other  hand,  where  the  direction  of  the  wind  is  quite 
variable,  it  may  be  stated  as  a  general  principle  that  windows  should  be  arranged  in 
sidewalls  and  monitors  so  that  there  will  be  approximately  equal  area  on  all  sides. 
Thus,  no  matter  what  the  wind 's  direction,  there  will  always  be  some  openings  directly 
exposed  to  the  pressure  force  of  the  wind,  and  others  opposed  to  a  suction  force,  and 
effective  movement  through  the  building  will  be  assured. 

11.  The  intensity  of  suction  or  the  vacuum  produced  by  the  jump  of  ^the  wind  is 
greatest  just  back  of  the  building  face.    The  area  of  suction  does  not  vary  with  the  wind 
velocity,  but  the  flow  due  to  suction  is  directly  proportional  to  wind  velocity. 

12.  Openings  much  larger  than  the  calculated  areas  are  sometimes  desirable,  especially 
when  changes  in  occupancy  are  possible,  or  to  provide  for  extremely  hot  days.    In  the 
former  case,  free  openings  should  be  located  at  the  level  of  occupancy  for  psychological 
reasons. 

13.  Special  consideration  should  be  given  to  the  possibility  of  sidewall  or  monitor 
windows  being  closed  on  account  of  weather  conditions.    Such  possibilities  favor  roof 
ventilators  and  specially  designed  stormproof  inlets. 

MEASUREMENT  OF  NATURAL  AIR  FLOW 

The  determination  of  the  performance  of  any  ventilating  system 
involves  measurements  which  are  not  easy  to  make.  The  difficulties  are 
increased  in  the  case  of  natural  ventilation,  since  the  motive  forces  and 
the  air  velocities  are  very  small.  The  measurements  necessary  for  giving 
the  capacity  of  a  system  are  (1)  velocity  of  the  wind,  (2)  velocity  of  the 
air  through  inlet  and  outlet  openings,  (3)  outdoor  air  temperature,  and 
(4)  average  indoor  air  temperature. 

Measuring  Wind  Velocity.  The  cup-type  of  anemometer  as  used  for 
Weather  Bureau  observations  is  sufficiently  accurate  for  this  measure- 
ment. Some  more  accurate  instruments  as  well  as  direct-reading  types 
have  been  developed  for  airport  service,  but  for  ventilation  work  it  is  the 
average  wind  velocity  over  a  long  period  which  determines  the  capacity  of 
the  system.  Hence  the  use  of  the  Weather  Bureau  instrument,  with  an 
observation  period  of  one  hour  or  more,  is  satisfactory.  If  observations 
of  wind  direction  are  required,  these  should  be  taken  by  observing  a 
sensitive  weather  vane  at  frequent  intervals  (about  every  5  minutes) 
during  the  same  period, 

Velocity  of  Air  Through  Openings.  The  vane  type  anemometer  is  the 
most  practical  instrument  for  this  measurement. 

86 


CHAPTER  4 — NATURAL  VENTILATION 


Use  a  small  (4  in.)  low-speed  anemometer,  and  correct  all  readings 
according  to  a  recent  calibration.  Mount  the  anemometer  in  a  strap  iron 
clamp  with  a  long  handle  for  convenience.  Divide  each  opening  into 
5  in.  squares  (by  string  or  wire)  and  hold  the  anemometer  in  the  center  of 
each  square  for  a  definite  period  of  from  15  to  30  seconds.  Record  the 
result  of  the  traverse  as  soon  as  completed  and  start  another  one  im- 
mediately. A  series  of  traverses  over  a  period  of  one  hour,  or  the  full 
period  covered  by  the  wind  velocity  observations  with  a  fairly  steady 
wind,  may  be  considered  a  satisfactory  test  for  that  wind  velocity.  It  is 
preferable  to  have  an  anemometer  observer  at  each  opening.  If  the 
opening  is  covered  by  a  grille  or  register,  use  the  proper  correction  factors 
(see  Chapter  40). 

Outdoor  Temperature.  It  is  easy  to  make  an  error  of  1  to  5  deg  in 
observing  ^the  outdoor _  air  temperature.  An  accurate  thermometer, 
calibrated  in  1  deg  divisions  should  be  used.  The  thermometer  should  be 
mounted  in  the  shade  at  about  mid-height  of  the  building  and  not  too 
near  the  building  wall  or  adjacent  to  an  air  outlet.  The  heat  from  a  wall 
or  roof  which  has  been  exposed  to  the  sun  is  easily  transmitted  to  a 
thermometer,  with  resulting  high  readings. 

Average  Indoor  Temperature.  It  is  important  to  note  that  the  capacity 
of  an  opening  (such  as  roof  ventilator)  does  not  depend  on  the  difference 
in  the  temperatures  measured  adjacent  to  the  opening.  It  depends 
rather  on  the  difference  between  the  average  temperature  of  the  column 
of  air  inside  the  building  and  that  outside.  Indoor  temperatures  should 
therefore  be  observed  at  various  heights  to  secure  a  good  average. 

DAIRY  BARN  VENTILATION4 

A  successful  barn  ventilating  system  is  one  which  continuously  supplies 
the  proper  amount  of  air  required  by  the  stock,  with  proper  distribution 
and  without  drafts,  and  one  which  removes  the  excessive  heat,  moisture, 
and  odors,  and  maintains  the  air  at  a  proper  temperature,  relative 
humidity,  and  degree  of  cleanliness. 

Barn  temperatures  below  freezing  and  above  80  F  affect  milk  produc- 
tion. Milk  producing  stock  should  be  kept  in  a  barn  temperature  be- 
tween 45  and  50  F.  Dry  stock,  at  reduced  feeding,  may  be  kept  in  a  barn 
5  to  10  deg  higher.  Calf  barns  are  generally  kept  at  60  F,  while  hospital 
and  maternity  barns  usually  have  a  temperature  of  60  F  or  somewhat 
higher. 

The  heat  produced  by  a  cow  of  an  average  weight  of  1000  Ib  may  be 
taken  as  3000  Btu  per  hour.  The  average  rate  of  moisture  production  by 
a  cow  giving  20  Ib  of  milk  per  day  is  15  Ib  of  water  per  day,  or  4375  grains 
per  hour.  To  set  a  standard  of  permissible  relative  humidity  for  cow 
barns  is  difficult.  For  45  F  an  average  relative  humidity  of  80  per  cent 
is  satisfactory,  with  85  per  cent  as  a  limit. 

Where  the  barn  volume  is  within  the  limit  that  can  be  heated  by  the 
stabled  animals,  the  air  supply  need  not  be  heated.  The  air  should  be 

*For  additional  information  on  this  subject  refer  to  Technical  Bulletin,  U.  S,  Department  of  Agriculture 
(1930),  by  M.  A.  R.  Kelley. 

Dairy  Barn  Ventilation,  by  F.  L.  Fairbanks  (A.S.H.V.E.  TRANSACTIONS,  Vol.  34,  1928). 

Cow  Barn  Ventilation,  by  Alfred  J.  Offner  (A^S.H.V.E.  Journal  Section,  Heating,  Piping  and  Air 
Conditioning.  January,  1933). 

87 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

supplied  through  or  near  the  ceiling.  It  is  better  to  have  the  exhaust 
openings  near  the  floor  as  larger  volumes  of  warm  air  are  then  held  in  the 
barn  and  there  is  better  temperature  control  with  less  likelihood  of  sudden 
change  in  barn  temperature. 

If  a  cow  weighs  1000  Ib  and  produces  3000  Btu  of  heat  per  hour,  and  if 
a  barn  for  the  cow  has  600  cu  ft  of  air  space  with  130  sq  ft  of  building 
exposure,  one  cow  will  require  2600  to  3550  cfh  of  ventilation,  depending 
on  the  temperature  zone  in  which  the  barn  is  located.  The  permissible 
heat  losses  through  the  structure,  based  on  one  cow  and  depending  on  the 
temperature  zone,  vary  between  0.043  and  0.066  Btu  per  hour  per  cu  ft 
of  barn  space,  and  0.197  to  0.305  Btu  per  hour  per  sq  ft  of  barn  exposure. 

GARAGE  VENTILATION-6 

On  account  of  the  hazards  resulting  from  carbon  monoxide  and  other 
physiologically  harmful  or  combustible  gases  or  vapors  in  garages,  the 
importance  of  proper  ventilation  of  these  buildings  cannot  be  over- 
emphasized. During  the  warm  months  of  the  year,  garages  are  usually 
ventilated  adequately  because  the  doors  and  windows  are  kept  open.  As 
cold  weather  sets  in,  more  and  more  of  the  ventilation  openings  are  closed 
and  consequently  on  extremely  cold  days  the  carbon  monoxide  concentra- 
tion runs  high. 

Many  garages  can  be  satisfactorily  ventilated  by  natural  means  par- 
ticularly during  the  mild  weather  when  doors  and  windows  can  be  kept 
open.  However,  the  A.S.H.V.E.  Code  for  Heating  and  Ventilating 
Garages,  adopted  in  1929,  states  that  natural  ventilation  may  be  em- 
ployed for  the  ventilation  of  storage  sections  where  it  is  practical  to 
maintain  open  windows  or  other  openings  at  all  times.  The  code  specifies 
that  such  openings  shall  be  distributed  as  uniformly  as  possible  in  at  least 
two  outside  walls,  and  that  the  total  area  of  such  openings  shall  be 
equivalent  to  at  least  5  per  cent  of  the  floor  area.  The  code  further  states 
that  where  it  is  impractical  to  operate  such  a  system  of  natural  ventilation, 
a  mechanical  system  shall  be  used  which  shall  provide  for  either  the  supply 
of  1  cu  ft  of  air  per  minute  from  out-of-doors  for  each  square  foot  of  floor 
area,  or  for  removing  the  same  amount  and  discharging  it  to  the  outside 
as  a  means  of  flushing  the  garage. 

Research 

Research  on  garage  ventilation  undertaken  by  the  A.S.H.V.E.  Com- 
mittee on  Research  at  Washington  University,  St.  Louis,  Mo.,  and  at  the 


*Code  for  Heating  and  Ventilating  Garages  (A.S.H.V.E.  TRANSACTIONS,  Vol.  35,  1929). 

Airation  Study  of  Garages,  by  W.  C.  Randall  and  L.  W.  Leonhard  (A.S.H.V.E.  TRANSACTIONS,  Vol.  36, 
1930). 

6Carbon  Monoxide  Concentration  in  Garages,  by  A.  S.  Langsdorf  and  R.  R,  Tucker  (A.S.H.V.E.  TRANS- 
ACTIONS, Vol.  36,  1930). 

Carbon  Monoxide  Distribution  in  Relation  to  the  Ventilation  of  an  Underground  Ramp  Garage,  by 
F.  C.  Houghten  and  Paul  McDermott  (A.S.H.V.E.  TRANSACTIONS,  Vol.  38,  1932). 

Carbon  Monoxide  Distribution  in  Relation  to  the  Ventilation  of  a  One-Floor  Garage,  by  F.  C.  Houghten 
and  Paul  McDermott  (A.S.H.V.E.  TRANSACTIONS,  Vol.  38,  1932). 

Carbon  Monoxide  Distribution  in  Relation  to  the  Heating  and  Ventilation  of  a  One-Floor  Garage,  by 
F.  C.  Houghten  and  Paul  McDermott  (A.S.H.V.E.  Journal  Section,  Healing,  Piping  and  Air  Conditioning, 
July,  1933). 

Carbon  Monoxide  Surveys  of  Two  Garages,  by  A.  H.  Sluss,  E.  K.  Campbell  and  Louis  M.  Farber 
(A.S.H.V.E.  Journal  Section,  Heating,  Piling  and  Air  Conditioning,  December,  1933). 


CHAPTER  4 — -NATURAL  VENTILATION 


University  of  Kansas,  Lawrence,  Kans.,  in  cooperation  with  the  A.S.H. 
V.E.  Research  Laboratory,  and  at  the  A.S.H.V.E.  Research  Laboratory 
has  resulted  in  authoritative  papers  on  the  subject. 

Some  of  the  conclusions  from  work  at  the  Laboratory  are  listed  below : 

1.  Upward  ventilation  results  in  a  lower  concentration  of  carbon  monoxide  at  the 
breathing  line  and  a  lower  temperature  above  the  breathing  line  than  does  downward 
ventilation,  for  the  same  rate  of  carbon  monoxide  production,  air  change  and  the  same 
temperature  at  the  30-in.  level. 

2.  A  lower  rate  of  air  change  and  a  smaller  heating  load  are  required  with  upward 
than  with  downward  ventilation. 

3.  In  the  average  case  upward  ventilation  results  in  a  lower  concentration  of  carbon 
monoxide  in  the  occupied  portion  of  a  garage  than  is  had  with  complete  mixing  of  the 
exhaust  gases  and  the  air  supplied.     However,  the  variations  in  concentration  from 
point  to  point,  together  with  the  possible  failure  of  the  advantages  of  upward  ventilation 
to  accrue,  suggest  the  basing  of  garage  ventilation  on  complete  mixing  and  an  air  change 
sufficient  to  dilute  the  exhaust  gases  to  the  allowable  concentration  of  carbon  monoxide. 

4.  The  rate  of  carbon  monoxide  production  by  an  idling  car  is  shown  to  vary  from 
25  to  50  cfh,  with  an  average  rate  of  35  cfh. 

5.  An  air  change  of  350,000  cfh  per  idling  car  is  required  to  keep  the  carbon  monoxide 
concentration  down  to  one  part  in  10,000  parts  of  air. 


PROBLEMS  IN  PRACTICE 

1  •  a.  What  means  are  available  for  the  ventilation  of  buildings? 

b.  What  precaution  is  necessary  in  combining  different  means  of  venti- 
lating? 

a.  Natural  forces,  such  as  winds  and  stack  effect,  and  mechanical  forces  furnished 
by  fans. 

b.  It  is  desirable  that  the  different  forces  used  be  not  in  opposition.    Their  actions  should 
be  mutually  helpful.    For  example,  a  simple  roof  opening  should  be  placed  in  the  region 
of  lowest  pressure  caused  by  a  prevailing  wind.    (See  Fig.  1.) 

2  •  a.  What  factors  are  important  in  the  location  and  control  of  ventilating 
openings? 

b.  What  types  of  ventilating  openings  are  best  suited  to  a  proper  distribu- 
tion of  the  air  supplied? 

a.  The  proper  distribution  of  air  as  required  by  the  occupants,  and  the  best  utilization 
of  natural  ventilating  forces.    The  general  rules  on  page  85  apply  particularly  to  these 
factors. 

b.  Windows  with  swinging  sash  and  openings  with  deflectors  may  be  used  to  direct  air 
to  the   points  desired. 

3  •  a.  What  is  the  best  location  for  ventilating  openings? 

b.  How  are  the  sizes  of  ventilating  openings  determined  for  proper  air 
supply? 

a.  Inlet  openings  should  be  low  and  facing  the  prevailing  winds  where  possible.    Outlet 
openings  should  be  high  and  on  the  side  opposite  the  prevailing  winds. 

b.  For  simple  openings  use  Formula  1: 

Q  =  EAV 
and  for  stacks  use  Formula  2: 

Q  =  9.4  A  V  H  (ti  -  fe) 

The  use  of  these  formulae  is  illustrated  in  Example  1  of  the  text  of  this  chapter.    Inlet 
and  outlet  areas  should  be  approximately  the  same  for  best  results. 

89 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

4  •  a.  What  are  the  advantages  of  roof  ventilators? 

h.  How  are  proper  sizes  determined  for  roof  ventilators? 

a.  Roof  ventilators  offer  the  best  utilization  of  the  inductive  force  of  the  wind,  and  they 
may  be  very  economically  fitted  with  built-in  fans  to  supply  the  necessary  circulation 
when  the  force  of  the  wind  is  not  sufficient. 

b.  Because  of  the  many  factors  affecting  the  flow  through  roof  ventilators  no  accurate 
formula  can  be  given.     It  is  usual  practice  to  make  the  combined  throat  area  of  all 
roof  ventilators  between  one-half  area  and  full  area  of  the  air  inlets  as  determined  by 
Formula  1. 

5  •  What  methods  of  control  are  used  in  ventilating  systems? 

Hand  control,  control  by  a  thermostat  located  in  the  ventilated  space  or  in  the  venti- 
lator, or  wind  velocity  control  designed  to  keep  the  air  discharge  constant  regardless 
of  wind  velocity. 

6  •  How  is  the  quantity  of  air  required  for  a  huilding  determined? 

Sufficient  air  must  be  supplied  to  carry  away  the  heat  and  impurities  generated  within  a 
building.  The  temperature  rise  and  concentration  of  impurities  in  the  exhaust  air  must 
be  held  within  specified  limits.  (See  Example  1.) 

7  •  What  measurements  are  necessary  to  determine  the  capacity  of  a  venti- 
lating system? 

Wind  velocity  and  air  velocities  through  openings,  determined  by  suitable  cup  anemo- 
meters; outdoor  air  temperatures,  measured  by  a  shaded  thermometer  not  near  objects 
heated  by  the  sun  or  near  exhaust  air  openings;  indoor  air  temperatures,  measured  at 
various  heights  to  secure  a  good  average. 

8  •  How  much  air  must  he  supplied  for  dissipating  the  heat  generated  in  a 
dairy  harn  housing  100  cows  if  the  outside  temperature  is  20  F  and  the  inside 
temperature  is  to  be  maintained  at  45  F? 

The  total  heat  generated  is  100  X  3000  =  300,000  Btu  per  hour  or  5,000  Btu  per 
minute.  Then  from  Formula  3, 

o-HV 

Q  ~  CD 

5000  X  13.5 
"~  0.24  X  (45  -  20) 
=  11,250  cu  ft  per  minute. 
This  amount  of  air  should  also  keep  down  humidity  and  odors. 

9  •  a.  What  precaution  is  necessary  in  the  ventilation  of  garages  using  natural 
ventilation? 

h.  How  much  window  area  is  required  for  a  garage  with  50  x  100  sq  ft  floor 
area  if  natural  ventilation  is  used? 

a.  The  carbon  monoxide  content  of  the  air  should  be  kept  below  1  part  in  10,000  and 
windows  should  be  kept  open  at  all  times. 

b.  The  window  area  should  aggregate  5  per  cent  of  the  floor  area. 

0.05  X  50  X  100  =  250  sq  ft  of  window  area. 
This  area  should  be  evenly  distributed  along  two  sides  of  the  building. 


90 


Chapter  5 

HEAT  TRANSMISSION   COEFFICIENTS 
AND  TABLES 

Heat  Transfer,  Calculations  for  Transmission  Losses,  Areas 
Where  Transmission  Losses  Occur,  Coefficients  of  Transmission, 
Table  of  Conductivities  and  Conductances,  Tables  of  Over-all 
Coefficients  of  Heat  Transfer  for  Typical  Building  Constructions 

*~r\O  maintain  specified  inside  temperature  conditions  and  determine 
JL    the  type  of  plant  required,  it  is  essential  to  know  the  transmission 
losses  of  a  structure  and  consider  them  in  conjunction  with  the  infiltration 
losses. 

Whenever  a  difference  in  temperature  exists  between  the  two  sides  of 
any  structural  material,  such  as  a  wall  or  roof  of  a  building,  a  transfer  of 
heat  takes  place  through  that  material.  When  the  inside  temperature  is 
the  higher,  heat  reaches  or  enters  the  inside  surface  of  the  wall  by  radia- 
tion and  convection,  because  the  air  and  objects  within  the  building  are 
always  warmer  than  the  inside  surface  of  the  wall  when  the  inside  air 
temperature  t  is  greater  than  the  outside  air  temperature  fe.  This  heat 
must  then  pass  through  the  material  of  the  wall  from  the  inside  to  the 
outside  surface  by  conduction,  and  is  finally  given  off  from  the  outside 
surface  by  radiation  and  convection,  provided,  of  course,  that  equilibrium 
has  been  established  and  all  four  temperatures  are  constant.  If  the  out- 
side temperature  is  the  higher,  the  reverse  process  takes  place. 

CALCULATIONS  FOR  TRANSMISSION  LOSSES 

The  calculations  for  heat  transmission  losses  are  made  by  multiplying 
the  area  A  in  square  feet  of  wall,  glass,  roof,  floor,  or  material  through 
which  the  loss  takes  place,  by  the  proper  coefficient  U  for  such  construc- 
tion or  material  and  by  the  temperature  difference  between  the  inside  air 
temperature  t  at  the  proper  level  (in  many  cases  not  the  breathing-line) 
and  the  outside  air  temperature  t0.  Therefore, 

fit  =  A  U  (t  -  O  (1) 

where 

Ht  =  Btu  per  hour  transmitted  through  the  material  of  the  wall,  glass,  roof  or 

floor. 

A  a*  area  in  square  feet  of  wall,  glass,  roof,  floor,  or  material,  taken  from  building 
plans  or  actually  measured.  (Use  the  net  inside  or  heated  surface  dimensions 
in  all  cases.) 

t  —  t0  =  temperature  difference  between  inside  and  outside  air,  in  which  t  must  always 
be  taken  at  the  proper  level.  Note  that  t  may  not  be  the  breathing-line 
temperature  in  all  cases* 

91 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

Heat  is  lost  from  a  building  by  transmission  through  all  of  those  sur- 
faces which  separate  heated  spaces  from  the  outside  air  or  from  unheated 
colder  spaces  within  the  building.  In  general,  five  kinds  of  surfaces  are 
involved:  (1)  outside  walls;  (2)  outside  glass;  (3)  inside  walls  or  parti- 
tions next  to  unheated  spaces;  (4)  ceilings  of  upper  floors,  either  below  a 
cold  attic  space  or  as  the  underside  of  a  roof  slab ;  and  (5)  floors  of  heated 
rooms  above  an  unheated  space. 

The  net  inside  wall  surface  is  usually  determined  by  reference  to  the 
scale  plans  and  elevations  of  the  building  concerned.  In  some  cases,  of 
course,  the  actual  building  may  have  to  be  measured.  The  total  area  of 
all  outside  openings  which  are  occupied  by  windows  and  doors  is  accurately 
measured  and  listed  as  glass.  The  glass  area  is  then  deducted  from  the 
total  outside  wall  area  for  each  room  and  the  difference  is  the  net  wall 
area.  If  there  are  no  partitions,  measure  from  the  inside  face  of  one  wall 
to  the  inside  face  of  the  next  wall.  The  areas  of  walls,  ceilings  and  floors 
next  to  cold  or  unheated  spaces  are  found,  of  course,  by  taking  the  inside 
dimensions  of  such  areas,  measured  on  the  heated  side. 

COEFFICIENTS  OF  TRANSMISSION 

The  coefficients  of  transmission  may  be  determined  by  means  of  the 
guarded  hot  box  or  the  Nicholls  heat  meter  described  in  Chapter  40,  or 
they  may  be  calculated  from  fundamental  constants.  Because  of  the 
unlimited  number  of  combinations  of  building  materials,  it  would  be 
impractical  to  attempt  to  determine  by  test  the  heat  transmission  co- 
efficient of  every  type  of  construction  in  use;  consequently,  in  most  cases 
it  is  advisable  to  calculate  these  coefficients. 

Symbols 

The  following  symbols  are  used  in  the  heat  transmission  formulae  in 
this  chapter: 

U  —  thermal  transmittance  or  over-all  coefficient  of  heat  transmission ;  the  amount  of 
heat  expressed  in  Btu  transmitted  in  one  hour  per  square  foot  of  the  wall,  floor,  roof  or 
ceiling  for  a  difference  in  temperature  of  1  deg  F  between  the  air  on  the  inside  and  that 
on  the  outside  of  the  wall,  floor,  roof  or  ceiling. 

k  =  thermal  conductivity;  the  amount  of  heat  expressed  in  Btu  transmitted  in  one 
hour  through  1  sq  ft  of  a  homogeneous  material  1  in.  thick  for  a  difference  in  temperature 
of  1  deg  F  between  the  two  surfaces  of  the  material.  The  conductivity  of  any  material 
depends  on  the  structure  of  the  material  and  its  density.  Heavy  or  dense  materials,  the 
weight  of  which  per  cubic  foot  is  high,  usually  transmit  more  heat  than  light  or  less  dense 
materials,  the  weight  of  which  per  cubic  foot  is  low. 

Ca  =  thermal  conductance  per  unit  area;  the  amount  of  heat  expressed  in  Btu  trans- 
mitted in  one  hour  through  1  sq  ft  of  a  non-homogeneous  material  for  the  thickness  or 
type  under  consideration  for  a  difference  in  temperature  of  1  deg  F  between  the  two 
surfaces  of  the  material.  Conductance  is  usually  used  to  designate  the  heat  transmitted 
through  such  heterogeneous  materials  as  plaster  board  and  hollow  clay  tile. 

f  —  film  or  surface  conductance;  the  amount  of  heat  expressed  in  Btu  transmitted  by 
radiation,  conduction  and  convection  from  a  surface  to  the  air  surrounding  it,  or  vice 
versa,  in  one  hour  per  square  foot  of  the  surface  for  a  difference  in  temperature  of  1  deg  F 
between  the  surface  and  the  surrounding  air.  To  differentiate  between  inside  and  outside 
wall  (or  floor,  roof  or  ceiling)  surfaces,  /i  is  used  to  designate  the  inside  film  or  surface 
conductance  and  /0  the  outside  film  or  surface  conductance. 

a  =  thermal  conductance  of  an  air  space;  the  amount  of  heat  expressed  in  Btu  trans- 
mitted by  radiation,  conduction  and  convection  in  one  hour  through  an  area  of  1  sq  ft  of 

92 


CHAPTER  5 — HEAT  TRANSMISSION  COEFFICIENTS  AND  TABLES 

an  air  space  for  a  temperature  difference  of  1  deg  F.  The  conductance  of  an  air  space 
depends  on  the  mean  absolute  temperature,  the  width,  the  position  and  the  character  of 
the  materials  enclosing  it. 

R  =  resjstance  or  resistivity  which  is  the  reciprocal  of  transmission,  conductance, 
or  conductivity,  i.e.: 

— —  =  over-all  or  air-to-air  resistance. 
—j-  =  internal  resistivity. 

K 

-~-  ~  internal  resistance. 
C-a 

-7-  —  film  or  surface  resistance. 
—  =  air-space  resistance. 

Fundamental  Formulae 

The  formula  of  the  over-all  coefficient  for  a  simple  wall  x  inches  thick  is: 

1 


U 


J_  +  JL  j_  _L 
A          k    +  /0 


and  for  a  compound  wall  of  several  materials  having  thicknesses  in  inches 
of  rci,  #a,  x3,  etc.,  the  coefficient  is: 


U 


In  the  case  of  air-space  construction,  an  air-space  coefficient  for  each 
air  space  must  be  inserted  in  either  Equation  2  or  3.  Thus  for  a  simple 
wall  with  one  air  space, 


U 


/0 


and  for  a  simple  wall  of  several  air  spaces  having  conductances  of 
a*,  a»,  etc.,  the  coefficient  is: 


U 


With  certain  special  forms  of  materials  which  have  irregular  air  spaces 
(such  as  hollow  tile)  or  are  otherwise  non-homogeneous,  it  is  necessary 
to  use  the  conductance  (Ca)  for  the  unit  construction,  in  which  case 

-r-  is  replaced  by  -~-. 

As  in  the  case  of  the  simple  wall,  /i  and  /0  are  always  the  inside  and 
outside  surface  coefficients  for  the  two  materials  in  contact  with  air.  If 

93 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


the  air  is  still  (no  wind),  then  for  the  same  material  f\  and/0  are  the  same, 
and/i  =  /0;  but  if  the  outside  air  is  in  motion,  then/0  is  always  greater 
than  /i  and  will  increase  as  the  wind  velocity  increases.  Values  for  fi  in 
still  and  moving  air  have  been  determined  for  various  building  materials 
at  the  University  of  Minnesota  under  a  cooperative  research  agreement 
with  the  Society1.  The  range  of  values  for  ordinary  building  materials  is 
comparatively  small  and  for  practical  purposes  may  be  assumed  constant 
for  either  still  air  or  any  given  wind  velocity,  particularly  in  view  of  the 
fact  that  the  surface  resistances  usually  comprise  only  a  small  part  of  the 
total  resistance  of  the  construction,  except  in  the  case  of  thin,  highly 
conductive  walls.  In  determining  basic  heat  transmission  values  for 
building  construction,  it  is  customary  to  use  that  value  of /0  which  will 
occur  when  a  15-mph  wind  blows  parallel  to  the  outer  surfaces  considered. 

TABLE  1.    CONDUCTANCES  OF  AIR  SPACES  a  AT  VARIOUS  MEAN  TEMPERATURES 


MEAN 
TUMP 
DBO  FAHK 

CONDUCTANCES  OF  AIR  SPACES  FOR  VARIOUS  WIDTHS  IN  INCHES 

0.128 

0.250 

0.364 

0.493 

0.713 

1.00 

1.500 

20 

2.300 

1.370 

1.180 

1.100 

1.040 

1.030 

1.022 

30 

2.385 

1.425 

1.234 

1.148 

1.080 

1.070 

1.065 

40 

2.470 

1.480 

1.288 

1.193 

1.125 

1.112 

1.105 

50 

2.560 

1.535 

1.340 

1.242 

1.168 

1.152 

1.149 

60 

2.650 

1.590 

1.390 

1.295 

1.210 

1.195 

1.188 

70 

2.730 

1.648 

1.440 

1.340 

1.250 

1.240 

1.228 

80 

2.819 

1.702 

1.492 

1.390 

1.295 

1.280 

1.270 

90 

2.908 

1.757 

1.547 

1.433 

1.340 

1.320 

1.310 

100 

2.990 

1.813 

1.600 

1.486 

1.380 

1.362 

1.350 

110 

3.078 

1.870 

1.650 

1.534 

1.425 

1.402 

1.392 

120 

3.167 

1.928 

1.700 

1.580 

1.467 

1.445 

1.435 

130 

3.250 

1.980 

1.750 

1.630 

1.510 

1.485 

1.475 

140 

3.340 

2.035 

1.800 

1.680 

1.550 

1.530 

1.519 

150 

3.425 

2.090 

1.852 

1.728 

1.592 

1.569 

1.559 

aThermal  Resistance  of  Air  Spaces,  by  F.  B.  Rowley  and  A.  B.  Algren  (A.S.H.V.E.  TRANSACTIONS, 
Vol.  35,  1929). 

The  conductances  of  air  spaces  at  various  mean  temperatures  and 
widths,  for  ordinary  building  materials,  are  given  in  Table  1.  These 
results  were  likewise  obtained  at  the  University  of  Minnesota  under  a  co- 
operative research  agreement  with  the  Society. 

Values  for  k  and  Ca,  the  conductivity  and  conductance  of  building  ma- 
terials and  insulations,  are  given  in  Table  2  as  taken  from  the  published 
values  of  various  investigators.  It  should  be  noted  that  values  of  -k  and 
Ca  as  well  as  of  U  are  dependent  on  the  mean  temperature,  and  it  is 
therefore  desirable  that  the  investigator  determine  heat-transmission 
values  under  conditions  approximating  those  existing  under  actual  con- 
ditions. Recommended  values  for  calculating  the  coefficients  of  trans- 
mission of  various  types  of  construction  are  marked  by  an  asterisk  in 
Table  2. 


^Surface  Conductances  as  Affected  by  Air  Velocity,  Temperature  and  Character  of  Surface,  by  F.  B. 
Rowley,  A.  B.  Algren  and  J.  L.  Blackshaw  (A.S.H.V.E.  TRANSACTIONS,  Vol.  36,  1930).  See  also  references 
at  end  of  chapter. 

94 


CHAPTER  5 — HEAT  TRANSMISSION  COEFFICIENTS  AND  TABLES 


TABLE  2.    CONDUCTIVITIES  (k)  AND  CONDUCTANCES  (Ca)  OF  BUILDING 
MATERIALS  AND  INSULATORS^ 

7V«r  coefficients  are  expressed  in  Btu  per  hour  per  square  foot  per  degree  Fahrenheit  per  1  in.  thickness, 
•unless  otherwise  indicated. 


I 
Material 

Description 

DENSITY 
(Le  PER  Cu  FT) 

If 

M 

CONDUCTIVITY  (fc) 

OR 

CONDUCTANCE  (Ca) 

CIS  8 

1s! 

3  3 

AUTHORITY 

MASONRY  MATERIALS 

Common                

5.00* 

0.20 

- 

Face 

9.20* 

0.11 

BRICKWORK. 

Damp  or  wet._     

5.00fr 

0.20 

(2) 

Typical. 

12.00* 

0.08 

UEMENT  MO   T       _ 

Typical                           ....          

110.0 

75 

5.20* 

0.19 

(3) 

f(^          fl        ,-Ttrge 

Typical  (8  in.) 

0.62f* 

1.61 

UINDBE  Dlt              

u       (12  in  j"                                  . 

„ 

0.511* 

1.96 

CONCRETE 

Typical  -  

12.00*> 

0.08 

1-2-4  mix.  -  
Various  ages  and  mixesd  

Cellular    -    -          

143.0 
40.0 

69 

75 

9.46 
11.35*0 
16.36 
1.06 

0.11 
0.94 

(4) 
(5) 

(3) 

50.0 

75 

1.44 

0.69 

(3) 

a. 

60  0 

75 

1.80 

0.56 

(3) 

a 

70.0 

75 

2.18 

0.46 

(3) 

Typical    gypsum    fiber    concrete,    87.5% 
gypsum  and  12  5%  "wood  chips 

51.2 

74 

1.66* 

0.60 

(4) 

CONCRETE  BLOCKS 

Special  concrete  made  with  an  aggregate 
of  hardened  clay  —  1-2-3  mix.  „  
Typical  (8  in  ) 

101.0 

70 

3.98 
l.OOf* 

0.25 
1.00 

(3) 

""      (12  in)                   

0.80f* 

1.25 

Special  concrete  block  made  with  an  aggre- 
gate of  hardened  clay  —  4  x  8  x  16  in., 
3  cores  18%  voids 

74  0 

0.66f 

1.51     - 

(X\ 

Special  concrete  block  made  with  an  aggre- 
gate of  hardened  clay—  8  x  8  x  16  in., 
4  cores  35%  voids 

74.5 

0.30f 

3.33 

(3) 

n 

Typical 

12.50* 

0.08 

STUCCO 

12.00* 

0.08 

TILE 

Typical  hollow  clay  (4  in.)  

i.oot* 

1.00 

(6  in.)" 

0.64t* 

1.57 

- 

(8  in  )« 

0.60J* 

1.67 

(10  in  )e 

0,58t* 

K72 

(1?  in  y 

0.40f* 

2.50 

(16  in)'  

0.31t* 

3.23 

• 

Hollow  clay  (2  in.)  M-in.  plaster  both  sides 
Hollow  clay  (4  in.)  H-in.  plaster  both  sides 
Hollow  clay  (6  in.)  ^in.  plaster  both  sides 
Hollow  gypsum  (4  in.)  

120.0 
127.0 
124.3 

110 
100 
105 

l.OOf 
0.60f 
0.47f 
0.46f 

1.00 
1.67 
2.13 
2.18 

(2) 

2) 
(2) 

51.8 

70 

1.66 

0.60 

(4) 

Solid  gypsum 

75.6 

76 

2.96 

0.34 

<±) 

TlLE  OR.  TBRRA.Z7O 

Typical  flooring 

12.00* 

0.08 

, 

AUTHORITIES: 

1U.  S.  Bureau  of  Standards,  tests  based  on  samples  submitted  by  manufacturers. 

2A.  C.  Willard,  L.  C.  Lichty,  and  L.  A,  Harding,  tests  conducted  at  the  University  of  Illinois. 

*J.  C.  Peebles,  tests  conducted  at  Armour  Institute  of  Technology,  based  on  samples  submitted  by  manufacturers. 

<F.  B.  Rowley,  tests  conducted  at  the  University  of  Minnesota. 

*A.S.H.V.E.  Research  Laboratory. 

6K  A.  AUcut,  tests  conducted  at  the  University  of  Toronto. 

''Lees  and  Chorlton. 

*Recommended  conductivities  and  conductances  far  computing  heat  transmission  coefficients. 

tFor  thickness  stated  or  used  on  construction,  not  per  1-in.  thickness. 

*For  additional  conductivity  data  see  Table  14,  Page  63, 19$4  A..S.R.E.  Data,  Book. 

^Recommended  value.   See  Heating,  Ventilating  and  Air  Conditioning,  by  Harding  and  Willard,  revised  edition,  1932. 

"One  air  cell  in  the  direction  of  heat  flow, 

<*See  A»SJB[.VJE.  Research  Paper,  Conductivity  of  Concrete,  by  F.  C.  Houghten  and  Carl  Gutberlet  (A.S.H.  V.E,  TRANS- 
ACTIONS, VoL  37,  1931). 

<The  6-in,,  8-in.,  and  10-in,  hollow  tile  figures  are  based  on  two  cells  in  the  direction  of  heat  flow.  The  124n.  hollow  tile 
is  based  on  three  cells  in  the  direction  of  heat  flow.  The  164n.  hollow  tile  consists  of  one  10-in,  and  one  6-m.  tile,  each  having 
two  cells  in  the  direction  of  heat  flow. 

-'Not  oompressed. 

Hoofing,  0,15-in.  thick  (1.34  Ib  per  sq  ft),  covered  witk  gravel  (0>83  ib  per  so;  ft),  combined  thickness  assumed  0.25. 

95 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


TABLE  2. 


CONDUCTIVITIES  (k)  AND  CONDUCTANCES  (Ca)  OF  BUILDING 
MATERIALS  AND  INSULATORS — Continued 


The  coefficients  are  expressed  in  Btu  per  hour  per  square  foot  per  degree  Fahrenheit  per  1  in,  thickness, 
unless  otherwise  indicated. 


Material 

i 
Description 

DENSITY 
(La  PER  Cu  FT) 

MEAN  TEMP. 
(DBQ  FAHR) 

CONDUCTIVITY  (k) 

OR 

CONDUCTANCE  (Ca) 

ci^g 
!1 

S 
o 

<! 

3) 
1) 
1) 
1) 
3) 
1) 

(1) 
CD 
(3) 
1) 
1) 
1) 
1) 
3) 
1) 
3) 

@ 

S 

(3) 

3) 
1) 

;i 
8 

1) 
3) 

ai 

(i) 

1 

3) 
1) 
(1) 

8 

i) 
i) 
i) 

(3) 

(3) 

INSULATION—  BLANKET 
OR  FLEXIBLE  TYPES 
FIBER...-  ,  „  

Typical  „_  _.. 

Chemically  treated  wood  fibers  held  between 
layers  of  strong  paper/  
Eel  grass  between  strong  paper  /_„  „... 

Flax  fibers  between  strong  paper/  _„ 
Hair  felt  between  layers  of  paper/  
Kapok  between  burlap  or  paper/.  

3.62 
4.60 
3.40 
4.90 
11.00 
1.00 

70 
90 
90 
90 
75 
90 

0.27* 

0.25 
0.26 
0.25 
0.28 
0.25 
0.24 

3.70 

4.00 
3.85 
4.00 
3.57 
4.00 
4.17 

INSULATION-SEMI- 
RIGID  TYPE 

Felted  cattle  hair/ 

13.00 
11.00 
12.10 
13.60 
7.80 
6.30 
6.10 
6.70 
10.00 
11.00 

90 
90 
70 
90 
90 
90 
90 
75 
90 
70 

0.26 
0.26 
0.30 
0.32 
0.28 
0.27 
0.26 
0.25 
0.37 
0.26 

3.84 
3.84 
3.33 
3.12 
3.57 
3.70 
3.85 
4.00 
2.70 
3.84 

Flax/  
Flax  and  rye/  

Felted  hair  and"  asbestos/  „„ 
75%  hair  and  25%  jute/  
50%  hair  and  50%  jute/  
Jute/ 

Felted  jute  and  asbestos/  
Compressed  peat  moss  

INSULATION—  LOOSE 
FILL  OR  BAT  TYPE 

Made  from  ceiba  fibers/  ,  . 

1.90 
1.60 

1.50 
9.40 

1.50 

4.20 
30.00 
24.00 
18.00 
12.00 
34.00 
26.00 
24.00 
19.80 
18.00 

T.TO 

21.00 
18.00 
14.00 
10.00 
14.50 
14.50 
11.50 

75 
75 

75 
103 

75 

72 
90 
90 
90 
90 
90 
90 
75 
90 
75 

90 
90 
90 
90 
90 
77 
75 
72 
86 
.36 

0.23 
0.24 

0.27 
0.27 

0.27 

0.24 
1.00 
0.77 
0.59 
0.44 
0.60 
0.52 
0.48* 
0.35 
0.34 
0.27* 
0.31 
0.30 
0.29 
0.28 
0.27* 
0.33 
0.38 
0.31 
1.04 
0.71 

4.35 
4.17 

3.70 
3.70 

3.70 

4.17 
1.00 
1.30 
1.69 
2.27 
1.67 
1.92 
2.08 
2.86 
2.94 
3.70 
3.22 
3.33 
3.45 
3.57 
3.70 
3.03 
2.63 
3.22 
0.96 
1.41 

GLASS  WOOL. 

Fibrous  material  made  from  dolomite  and 

silina.    r     ._-,  1L.    n 

Fibrous  material  made  frnm  slag,     „, 

Fibrous  material  25  to  30  microns  in  dia- 
meter, made  from  virgin  bottle  glass  
Made  from  combined  silicate  of  lime  and 

{ihltninf*-                  ,          .-.,L,-r  r-                         ,r         .      ,L     ,                             lr               ,, 

GEANTJLAR_ 
N 

GYPSUM,  ,  „, 

MINERAL  WOOL™.  ,~. 

RKGRANTTI.ATEI>  CORK 

Cellular,  dry  

«            a  ~  * 

Flaked,  dry  and  fluffy/  

«       «      «        « 

U               it           tt               « 

All  forms,  typical  

About  2is-in.  particles 

ROCK  WOOL 

Fibrous  material  made  from  rock  

«                 u            «          «          u        

Rock  wool  with  a  binding  agent  
Rock  wool  with  flax,  straw  pulp,  and  binder 
Rock  wool  with  vegetable  fibers  _. 

SAWDUST  -    . 

Ordinary^  .  „  „  .   ... 

SHAVINGS  

Ordinary-  



INSULATION-RIGID 

CORKBOAED      

Typical— 

0.30* 
0.34 
0.30 
0.27 
0.25 
0.32 
0.33* 
0.36 

0,38 

3.33 
2.94 
3.33 
3.70 
4.00 
3.12 
3.03 
2.78 

2.63 

FIBER.    .  .  .    J 

No  added  binder  .,  

«        «            u. 

14.00 
10.60 
7.00 
5.40 
14.50 

20.00 
25.00 

90 
90 
90 
90 
90 

70 

75 

u.        tt             tt 
«        «            a.      

Asphaltic  binder  ,  
Typical  , 

Made  from  chemically  treated  wood  fiber  
Made  from  chemically  treated  wood  and 
vegetable  fibers  ,.  ^  ...  „  

For  notes  see  Page  95. 


96 


CHAPTER  5 — HEAT  TRANSMISSION  COEFFICIENTS  AND  TABLES 


TABLE  2.    CONDUCTIVITIES  (k)  AND  CONDUCTANCES  (Ca)  OF  BUILDING 
MATERIALS  AND  INSULATORS — Continued 

The  coefficients  are  expressed  in  Btu  per  hour  per  square  foot  per  degree  Fahrenheit  per  1  in.  thickness, 
unless  otherwise  indicated. 


Material 

Description 

DENSITY 
(Lu  PER  Cu  FT) 

MEAN  TEMP. 
(DEO  FAHR) 

CONDUCTIVITY  (k) 

OR 

CONDUCTANCE  (C  ) 

^B 

£Sfj 

M            g 

«  £ 

-< 

INSULATION—  RIGID 
—  Continued 
FIBER 

Made  from  corn  stalks  - 

15.00 

71 

0.33 

3  03 

M) 

"      u    exploded  wood  fiber  
"      "    hard  wood  fibers  
Insulating  plaster  9/10-in.  thick  applied  to 
%-in.  plaster  board  base  
Made  from  licorice  roots.,  
Made  from  85%  magnesia  and  15%  asbestos 
Made  from  shredded  wood  and  cement  
*      "    sugar  cane  fiber-           

17.90 
15.20 

54.00 
16.10 
19.30 
24.20 
13.50 

78 
70 

75 
81 
86 
72 
70 

0.32 
0.32 

1.07f 
0.34 
0.51 
0.46 
0.33 

3.12 
3.12 

0.93 
2.94 
1.96 
2.17 
3.03 

(4) 
(3) 

(3) 
(3) 
(1) 
(3) 
H) 

Sugar  cane  fiber  insulation  blocks  encased  in 
asphalt  membrane  
Made  from  wheat  straw  _  _  ~ 
"           wood  fiber~—  

13.80 
17.00 
15.90 
15.00 

70 
68 
72 
70 

0-30 
0.33 
0.33 
0.33 

3.33 
3.03 
3.03 
3.03 

(3) 
(3) 
3) 

31 

uu                      «    _.... 

T.s"o 

15.20 

52 
72 

0.33 
0.29 
0.33 

3.03 

3.45 
3.03 

6) 
3) 

nt 

*    _  

16.90 

90 

0.34 

2.94 

(i) 

BUILDING  BOARDS 
ASBESTOS—  —  - 

Compressed  cement  and  asbestos  sheets  
Corrugated  asbestos  board  ...    _  „  

123.00 
20.40 

86 
110 

2.70 
0.48 

0.37 
2.08 

(i) 
(?) 

GTPSTTML 

Pressed  asbestos  mill  board  
Sheet  asbestos  

Gypsum  between  layers  of  heavy  paper 

60.50 
48.30 
62  80 

86 
110 
70 

0.84 
0.29 
1.41 

1.19 

3.45 
0  71 

CD 

(2) 

H) 

PLASTER  BOARD 

Rigid,  gypsum  between  layers  of  heavy 
paper  (J4-in.  thick)  
Gypsum  mixed  with  sawdust  between  layers 
of  heavy  paper  (0.39-in.  thick)  
(3*3  "L)--   ,                                 .  . 

53.50 
60.70 

90 
90 

2.60f 

3.60f 
3.73J* 

0.38 

0.28 
0.27 

(1) 
CD 

(lx£  m>j  



_.. 

2.82f* 

0.35 

_.. 

ROOFING  CONSTRUCTION 
ROOFING 

Asphalt,  composition  or  prepared  

70.00 

75 

6.50P 

0.15 

m 

SHINGLES.   .„  _ 

Biult  up  —  %-in.  thick  
Built  up,  bitumen  and  felt,  gravel  or  slag 
surfaced"  „  
Plaster  board,  gypsum  fiber  concrete  and 
3-ply  roof  covering,  _ 
Agbftstos 

52.40 
65.00 

76 
75 

3.53f* 
1.33t 

0.581 

6-OOf* 

0.28 
0.75 

1.72 
0.17 

(2) 

(4) 
(3) 

Asphalt, 

70.00 

75 

6.50J* 

0.15 

f3) 

Sla'te 
Wood  

201.00 

10.37* 
1.28f 

0.10 
0.78 

(7) 

PLASTERING  MATERIALS 
PxAB-rmB.-™.^  

CJfimfint                                           ,  -   , 

8.00 

0.13 

(2) 

Gypsum,  typical  ,  

Thickness  %  in 

73 

3.30* 
8.80t 

0.30 
0.11 

(T) 

METAL  LATH  AND  PLASTER  
WOOD  LATH  AND  PLASTER  

Total  thickness  %  in  
H-ifc-  plaster,  total  thickness  %  in  —  .  



70 

4.40f* 
2.50J* 

0.23 
0.40 

(4) 

BUILDING 
CONSTRUCTIONS 
FRAME  _ 

1-in.  fir  sheathing  and  building  paper_  
1-in.   fir  sheathing,    building   paper,   and 
yellow  pine  lap  aiding.,  ^  ,  „.  ^  _.  

, 

30 
20 

0.71t* 

o.sot* 

1.41 
2.00 

(4) 
(4) 

FLOORING 

1-in.  fir  sheathing,  building  paper  and  stucco 
Pine  lap  siding  and  building  paper  —  aiding 
4  in.  wide  
Yellow  pine  lap  siding  
Maple  —  across  grain 

40".00 

20 
16 
75 

0.82f* 

0.85f* 
1.28f* 
1.20 

1,22 

1.18 
0.78 
0.83 

(4) 

(4) 

(7) 

Battleship  linoleum  CJ^~i^  ) 

1.36f* 

0.74 

For  notes  see  Page  95. 


97 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


TABLE  2.    CONDUCTIVITIES  (k)  AND  CONDUCTANCES  (Ca)  OF  BUILDING 
MATERIALS  AND  INSULATORS — Continued 

The  coefficients  are  expressed  in  Btu  per  hour  per  square  foot  per  degree  Fahrenheit  per  1  in.  thickness, 
unless  otherwise  indicated. 


J 
Material 

Description 

DENSITY 
(Ln  PER  Cu  FT) 

1 

CONDUCTIVITT  (k) 

OH 

CONDUCTANCE  (Ca) 

3s 

§    i 

<§  s 

AUTHORITY 

AIR  SPACE  AND  SURFACE 
COEFFICIENTS 
Am  ppA.rRp 

•   •  •     •  • 
Over  %-in,  faced  with  ordinary  building 

materials                                               •-   - 

40 

not* 

0.91 

(4) 

_                 ~ 

Still  air  (/i)            .             -  

1.65f* 

0.61 

(4) 

SURFACES,                        

IS  mph  —  (/o)                                 

6.00f* 

0.17 

(4) 

j,                                  Tj                             .                          , 

Still  air  (/i)                   

60 

i.iat 

0.85 

AIR  SPACESTACED  WITH 
BRIGHT  ALUMINUM 
FOIL 

Air  space,  faced  one  side  with,  bright  alumi- 
num foil,  over  iNt-in.  wide  ..  ... 
Air  space,  faced  one  side  with  bright  alumi- 
num foil,  5£-in.  wide  .  - 
Air  space,  faced  both  sides  with  bright 
aluminum  foil  over  3^-in.  wide  

50 
50 
50 

0.46f* 
0.62t 
0.41  f* 

2.17 
1.61 

2.44 

(4) 
(4) 

(4) 

Air  space,  faced  both"  sides  with  bright 
j\IiiTTunum  foil  5^-in  'widft 

50 

O.S7f 

1.75 

(4) 

Air  space  divided  'in  two  with  single  curtain 
of  bright  aluminum  foil  (both  sides  bright) 
Each  space  over  /^-in.  wide         -    

50 

0.23f* 

4.35 

(4) 

Each  space  i^-in.  wide  .  .~  .  .  

Air  space  with  multiple  curtains  of  bright 
aluminum   foil,    bright   on   both   sides, 
curtains    more    than    %-in.    apart,    in 
standard  construction  

2  curtains  forming  3  spaces 

50 
50 

O.Slf 
O.lSf* 

3.23 
6.67 

(4) 
(4) 

3  curtains  forming  4  spjujss 

50 

O.llf* 

9.09 

(4) 

4  curtains  forming  5  spaces  

50 

0.09t* 

11.11 

(4) 

WOODS  (Across  Grain) 
BALSA 

20.0 

90 

0.58 

1.72 

(1) 

8.8 

90 

0.38 

2.63 

1) 

7.3 

90 

0.33 

3.03 

1) 

CALIPORNTA^Tl'BTyWOOr* 

0%  moisture       -    

22.0 

75 

0.66 

1.53 

4) 

Q%          u                                              

28.0 

75 

0.70 

1.43 

4) 

8%        «     

22.0 

75 

0.70 

1.43 

4) 

8%         " 

28.0 

75 

0.75 

1.33 

4) 

16%        u 

22.0 

75 

0.74 

1.35 

4) 

16%        «           

28.0 

75 

0.80 

1.25 

(4) 

CYPBBSP 

28.7 

86 

0.67 

1.49 

(1) 

0%  moisture 

26.0 

75 

0.61 

1.64 

(4) 

1      .    .                           n,,..^ 

0%        u 

34  0 

75 

0.67 

1.49 

f4) 

8%        " 

26.0 

75 

0.66 

1.52 

(4) 

8%        " 

34.0 

75 

0.75 

1.33 

(4) 

169'        " 

26.0 

75 

0.76 

1.32 

4) 

16^         u 

34.0 

75 

0.82 

1.22 

4) 

EASTERN  HEMLOCK 

0%  moisture                         .             

22.0 

75 

0.60 

1.67 

4) 

30.0 

75 

0.76 

1.32 

4} 

gm        « 

22.0 

75 

0.63 

1.59 

4) 

^1    ;  -  

30.0 
22.0 

75 
75 

0.81 
0.67 

1.23 
1.49 

(4) 
(4) 

16%        " 

30.0 

75 

0.85 

1.18 

(4) 

HARD  MAPLE  ' 

0%  moisture 

40.0 

75 

1.01 

0.99 

(4) 

"iv""-"*—  """  "  "* 

46.0 

75 

1.05 

0.95 

4) 

gw        « 

40.0 

75 

1.08 

0.93 

4) 

om             « 

46.0 

75 

1.13 

0.89 

4) 

16*7        * 

40.0 

75 

1.15 

0.87 

4)' 

16%    "  

46.0 

75 

1.21 

0.83 

4) 

For  notes  see  Page  95. 


CHAPTER  5 — HEAT  TRANSMISSION  COEFFICIENTS  AND  TABLES 


TABLE  2.    CONDUCTIVITIES  (k)  AND  CONDUCTANCES  (Ca)  OF  BUILDING 
MATERIALS  AND  INSULATORS — Continued 

The  ccejfir.ients  are  expressed  in  Btu  per  hour  per  square  foot  per  degree  Fahrenheit  per  1  in.  thickness, 
unless  other-wise  indicated. 


Material 

Description 

DENSITY 
(La  PER  Cu  FT) 

MEAN  TKMP. 
(DEO  FARE) 

CONDUCTIVITY  (k) 
on 
CONDUCTANCE  (C&) 

Is! 

S  & 

AUTHOHITY 

WOODS—  Continued 
LONGLEAF  YELLOW  PINE    _ 

0%  moisture       

30.0 

75 

0.76 

1.32 

(4) 

0% 

40.0 

75 

0.86 

1.16 

(4) 

8% 

30.0 

75 

0.83 

1.21 

C4> 

^% 

40  0 

75 

0.95 

1-05 

M) 

16% 

30.0 

75 

0.89 

1.12 

(4) 

l^  or 

40.0 

75 

1.03 

0.97 

(4) 

MAHOGANY 

34.3 

86 

0.90 

1.11 

f1> 

44  3 

86 

1  10 

0.91 

0) 

\^APTi1?5  ^R  OATC 

1.15* 

0.87 

NORWAY  PINE,™  r  ,  ,   »T  

§mm"ntnre 

22.0 

75 

0.62 

1.61 

(4) 

32.0 

75 

0.74 

1.35 

(4) 

22.0 

75 

0.68 

1.47 

ffl 

32.0 

75- 

0.83 

1.21 

(4) 

•\ffi7 

22.0 

75 

0.74 

1.35 

(4) 

\(\°7 

32.0 

75 

0.91 

1.10 

{4 

RED  CYPKBSS- 

n°7  moisture 

22.0 

75 

0.67 

1.49 

4 

Qcr 

32  0 

75 

0  79 

1.27 

4 

gcr 

22.0 

75 

0.71 

1.41 

4 

8% 

32.0 

75 

0.84 

1.19 

4 

1^% 

22.0 

75 

0.74 

1.35 

4) 

1^% 

32.0 

75 

0.90 

1.11 

4) 

P,*m  OAW 

0%  moisture 

38.0 

75 

0.98 

1.02 

4) 

48.0 

75 

1.18 

0.85 

4) 

8% 

38.0 

75 

1.03 

0.97 

4) 

jjor 

48.0 

75 

1.24 

0.81 

4) 

•j^O/ 

38.0 

75 

1.07 

0.94 

4) 

•Jri^ 

48.0 

75 

1.29 

0.78 

4) 

SJHOSTT..TAV  YfeL^OW  PfNTB  .  . 

O^7"  Tpoi^ttire 

26.0 

75 

0.74 

1.35 

4) 

n^ 

36.0 

75 

0.91 

1.10 

4) 

8% 

26.0 

75 

0.79 

1.27 

4) 

?% 

36.0 

75 

0.97 

1.03 

(4 

16% 

26.0 

75 

0.84 

1.19 

(4 

16% 

36.0 

75 

1.04 

0.96 

SOFT  Er,v 

n%  mois  ure  lr                  , 

28.0 

75 

0.73 

1.37 

(4 

n% 

34.0 

75 

0.88 

1.14 

4 

28.0 

75 

0.77 

.30 

4 

^ttr 

34.0 

75 

0.93 

.08 

4 

•j^ttr 

28.0 

75 

0.81 

.24 

4 

16%                

34.0 

75 

0.97 

.03 

4 

0%  moisture  

36.0 
42.0 

75 
75 

0.95 

.05 

4) 

fftf 

36.0 

75 

0.96 

.04 

4} 

8% 

42.0 

75 

1.02 

.98 

4) 

169^ 

36.0 

75 

1.01 

.99 

4) 

16% 

42.0 

75 

1.09 

.92 

4> 

SiftJAR  PINE  

0%  mois  rrrA  „,   ,         

22.0 

75 

0.54 

.85 

28.0 

75 

0.64 

.56 

4 

$Of 

22.0 

75 

0.59 

.70 

4 

9P7 

28.0 

75 

0.71 

.41 

4 

\fp? 

22.0 

75 

0.65 

.54 

4 

1^% 

28.0 

75 

0.78 

.28 

4 

VTWOTWTA,  Pr^ 

34.3 

86 

0.96 

1) 

Www  COAST  HKMKK^ 

0%  moisture...  .L  ^   ^    ,  

22.0 

75 

0.68 

.47 

4> 

0% 

30.0 

75 

0.79 

.27 

4> 

W? 

22.0 

75 

0.73 

.37 

4> 

%°7 

30.0 

75 

0.85 

.18 

4) 

\f\°7 

22.0 

75 

0.78 

.28 

4> 

\& 

30.0 

75 

0.91 

.10 

4) 

WBTTTB  PjN^.rnirJ1  ._  „  _ 

' 

31.2 

86 

0.78 

.28 

Yur.T.nw.  Prism          -,„,„,. 

1.00 

.00 

D 

YELLOW  PINK  OR  "Prp    JU-J 

0.80* 

1.25 

For  notes  see  Page  95. 


99 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 
TABLE  3.   COEFFICIENTS  OF  TRANSMISSION  ( U)  OF  MASONRY  WALLS<J 

Coefficients  are  expressed  in  Btu  per  hour  per  square  foot  per  degree 
Fahrenheit  difference  in  temperature  between  the  air  on  the  two  sides, 
and  are  based  on  a  wind  velocity  of  15  mph. 


THICKNESS 

r 

rYp 

[CAL 

OP 

WALL 

CONSTRUCTION 

TYPE  OF  WALL 

MASONRY 

No. 

(INCHES) 

I 

c 

£ 
t 

^ 

a 

3C 

•^»*, 

53Sp>j 

53 
383 

H^? 

^ 

,/TUCCO\ 

SoUd  Brick 

Based  on  4-in.  face  brick  and  the  remainder 
common  brick. 

B 
12 
16 

1 

2 
3 

—-  * 

^ 

m 

SS= 

y 

Hollow  Tile 

T 

^. 

^tSr 

^^ 

Stucco  Exterior  Finish. 

The  8-in.  and  10-in.  tile  figures  are  based  on 

8 

4 

two  cells  in  the  direction  of  flow  of  heat.    The 

10 

-     5 

12-in.  tile  is  based  on  three  cells  in  the  direc- 

12 

6 

tion  of  flow  of  heat.    The  16-in.  tile  consists 

16 

7 

Li 

*+, 

^^ 

*^^ 

^ 

of  one  10-in.  tile  and  one  6-in.  tile  each  having 
two  cells  in  the  direction  of  heat  flow. 

I 

-=±: 

t 

^ 

—  ^^. 

^ 

0 

& 

S 

tp. 

K 

T  ^ 

& 

12 

9 

:       1 

%.. 

£*>?f 

j 

Limestone  or  Sandstone 

16 

10 

1 

1 

-ji 

•*-*^ 

\> 

? 

24 

11 

^ 

~*  i. 

Concrete 

6 

12 

..  ', 

w  « 

o. 

»   - 

These  figures  may  be  used  with  sufficient 

10 

13 

accuracy   for   concrete    walls    with    stucco 

16 

14 

-;' 

o  ' 
—  il> 

1^  . 

£ 

exterior  finish. 

20 

15 

f=\ 

jg 

g 

g^ 

n 

Hollow  Cinder  Blocks 

S 

16 

Based  on  one  air  cell  in  direction  of  heat  flow. 

12 

17 

r 

Hollow  Concrete  Blocks 

B 

IS 

^ 

^ 
****** 

La, 

•~. 

t 

53S 

—». 

^ 

/ 

I 

Based  on  one  air  cell  in  direction  of  heat  flow. 

12 

19 

"Computed  from  factors  marked  by  *  in  Table  2. 
6  Based  on  the  actual  thickness  of  2-in.  furring  strips. 

100 


CHAPTER  5 — HEAT  TRANSMISSION  COEFFICIENTS  AND  TABLES 


INTERIOH  FINISH 


UNINSULATED  WA.LLS 


INSTTLA.TED  WALLS 


1 

1 

c 

§ 

s 

H 

3  is 

-1 

||.l 

ja 

i 

* 

3 

•f 

1 

S 

2,3  ^ 

"cS  g*O 

li 

1^ 

-s 

"S 

^ 

•S 

T3 

•*" 

^7 

5&01 

J»  2 

•£&"•!>. 

in  walla  —  no  interior  fir 

ister  (^  in.)  on  walls 

«2 

k 
1 

c 
o 

ister  (%  in.)  on  metal  1 

•o, 
J!i" 

corated  building  boar 
hout  plaster  —  furred 

5 
•g 

o 

-2  S 

a 
o 

&l 

It 

tflter  (li  in.)  on  corkboa 
in  cement  mortar  (l/i  \ 

ster  (%  in.)  on  metal  la 
furring  strips  —  furred  i 
in.  wide)  faced  one 
ght  aluminum  foil 

15 

§f| 

ister  (%  in.)  on  metal  la 
furring  strips  (2  in. 
ulation  (*<£  m.)  betwt 
ipa  (one  sur  space) 

5 

fi 

s 

g 

pui 

Q'S 

E^ 

Sci 

ei 

SS^S 

las 

ssJ-g 

A 

B 

c 

D 

E 

F 

G 

H 

i 

J 

K 

L 

0.50 

0.46 

0.30 

0.32 

0.30 

0.23 

0.22 

0.16 

0.14 

0.23 

0.12 

0.20 

0.36 

0.34 

0.24 

0.25 

0.24 

0.19 

0.19 

0.14 

0.12 

0.19 

0.11 

0.17 

0.28 

0.27 

0.20 

0.21 

0.20 

0.17 

0.16 

0.13 

0.11 

0.17 

0.10 

0.15 

0.40 
0.39 

0.37 
0.37 

0.26 
0.26 

0.27 
0.27 

0.26 
0.26 

0.20 
0.20 

0.20 
0.19 

0.15 
0.15 

0.13 
0.13 

0.20 
0.20 

0.11 
0.11 

0.18 
0.18 

0.30 

0.29 

0.22 

0.22 

0.22 

0.17 

0.17 

0.13 

0.12 

0.17 

0.10 

0.16 

0.25 

0.24 

0.19 

0.19 

0.19 

0.15 

0.15 

0.12 

0.11 

0.15 

0.097 

0.14 

0.71 

0.64 

0.37 

0.39 

0.37 

0.26 

0.25 

0.18 

0.15 

0.26 

0.13 

0.23 

0.58 

0.53 

0.33 

0.34 

0.33 

0.24 

0.23 

0.17 

0.14 

0.24 

0.13 

0.21 

0.49 

0.45 

0.30 

0.31 

0.30 

0.22 

0.22 

0.16 

0.14 

0.22 

0.12 

0.20 

0.37 

0.35 

0.25 

0.26 

0.25 

0.20 

0.19 

0.15 

0.13 

0.20 

0.11 

0.18 

0.79 

0.70 

0.39 

0.42 

0.39 

0.27 

0.26 

0.19 

0.16 

0.27 

0.13 

0.23 

0.62 

0.57 

0.34 

0.37 

0.34 

0.25 

0.24 

0.18 

0.15 

0.25 

0.13 

0.22 

0.48 

0.44 

0.29 

0.31 

0.29 

0.22 

0.21 

0.16 

0.14 

0.22 

0.12 

0.20 

0.41 

0.39 

0.27 

0.28 

0.27 

0.21 

0.20 

0.15 

0.13 

0.21 

0.12 

0.18 

0.42 

0.39 

0.27 

0.28 

0.27 

0.21 

0.20 

0.16 

0.13 

0.21 

0.12 

0.19 

0.37 

0.35 

0.25 

0.26 

0.25 

0.19 

0.19 

0.15 

0.13 

0.19 

0.11 

0.17 

0.56 

0.52 

0.32 

0.34 

0.32 

0.24 

0.23 

0.17 

0.14 

0.24 

0.12 

0.21 

0.49 

0.46 

0.30 

0.32 

0.30 

0.23 

0.22 

0.16 

0.14 

0.23 

0.12 

0.20 

«A  waterproof  membrane  should  be  provided  between  the  outer  material  and  the  insulation  fill  to 
prevent  possible  wetting  by  absorption  and  a  subsequent  lowering  of  efficiency. 


101 


AMERICAN  SOCIETY-  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

TA.BLE  4.    COEFFICIENTS  OF  TRANSMISSION  (U)  OF  MASONRY  WALLS 
WITH  VARIOUS  TYPES  OF  VENEERS* 


Coefficients  are  expressed  in  Btu  -per  hour  per  square  foot  per  degree 
Fahrenheit  difference  in  temperature  between  the  air  on  the  two  sides, 
and  are  based  on  a  wind  velocity  of  15  mph. 


TYPICAL 
CONSTRUCTION 


TYPE  OF  WALL 


FACING 


BACKING 


WALL 
No. 


4  in.  Brick  Veneer^ 


6  in. 

Sin. 
10  in. 
12  in. 


Hollow  Tile* 


4  in.  Brick  Veneer* 


Gin. 
10  in.    Concrete 

16  in. 


4  in.  Brick  Veneer'' 


8  in. 
12  in. 


Cinder  Blocks* 


4  in.  Brick  Veneer'' 


Sin. 
12  in. 


Concrete  Blocks* 


4  in.  Cut-Stone  Veneer* 


8  in. 
12  in.     Common  Brick 

16  in. 


4  in.  Cut-Stone  Veneer<* 


6  in. 

10  in 
12  in. 


Hollow  Tile- 


4  in.  Cut-Stone  Veneerd 


6  in, 
10  in.    Concrete 

16  in. 


20 
21 
22 
23 


24 
25 
26 


27 
28 


29 
30 


31 
32 
33 


34 
35 
36 
37 


38 
39 
40 


flComputed  from  factors  marked  by  *  in  Table  2. 
6  Based  on  the  actual  thickness  of  2-in,  furring  strips. 

*The  6-fn.,  8-in.  and  10-in.,tile  figures  are  based  on  two  cells  in  the  direction  of  heat  flow.    The  12-in. 
tile  is  based  on  three  cells  in  the  direction  of  heat  flow. 

102 


CHAPTER  5 — HEAT  TRANSMISSION  COEFFICIENTS  AND  TABLES 


INTERIOR  FINISH 


UNINSULATED  WALLS 


INSULATED  WALLS 


f! 

ii 

g 

g 

c 

•S|3 

"81 

•35  S 

J.s 

3 

£ 

§ 

If* 

3.2 

111 

JS, 

"8 

J, 

3 

S^ 

8 

X 

s 

III 

*"§ 

jjt  fl 

•1 

g 

J5 

J 

-0 

S 

X 

q    en 

11 

jlla  1 

s 

JS 

'-• 

T 

-2 

*•! 

•c 

•S1 

"^  £;  o 

His' 

Plain  walls  —  no  inter 

Plaster  (^  in.)  on  wa 

§ 

Plaster  (^  in.)  on  m 

g 

!f 

-2  5 

li 

No  plaster  —  decorate 
ing  board  interior  i 
furred 

a 
o 

^5 

^1 

&? 

o 

sl 

I! 

Plaster  on  corkboard 
cement  mortar  (H  in 

Plaster  on  metal  lath 
to  furring  strips—  fui 
J£:in.  wide)  faced 
bright  aluminum  foil 

Plaster  (%  in.)  on  me 
to  furring  strips  (2  i 
fill  (1%  in.b)/ 

Plaster  (%  in.)  on  me 
to  furring  strips  ( 
insulation  (^  m.) 
strips  (one  air  space) 

A 

B 

c 

D 

E 

F 

G 

H 

I 

J 

K 

L 

0.36 

0.34 

0.24 

0.25 

0.24 

0.19 

0.19 

0.16 

0.13 

0.19 

0.11 

0.17 

0.34 

0.33 

0.24 

0.25 

0.24 

0.19 

0.18 

0.14 

0.12 

0.19 

0.11 

0.17 

0.34 

0.32 

0.23 

0.24 

0.23 

0.19 

0.18 

0.14 

0.12 

0.19 

0.11 

0.17 

0.27 

0.26 

0.20 

0.21 

0.20 

0.16 

0.16 

0.13 

0.11 

0.16 

0.10 

0.15 

0.57 

0.53 

0.33 

0.35 

0.33 

0.24 

0.23 

0.17 

0.14 

0.24 

0.13 

0.21 

0.48 

0.45 

0.30 

0.31 

0.30 

0.22 

0.22 

0.16 

0.14 

0.22 

0.12 

0.20 

0.39 

0.37 

0.26 

0.27 

0.26 

0.20 

0.19 

0.15 

0.13 

0.20 

0.11 

0.18 

0.35 
0.31 

0.33 
0.30 

0.24 
0.22 

0.25 
0.23 

0.24 
0.22 

0.19 
0.18 

0.18 
0.17 

0.14 
0.14 

0.12 
0.12 

0.19 
0.18 

0.11 
0.11 

0.17 
0.16 

0.44 

0.42 

0.28 

0.30 

0.28 

0.21 

0.21 

0.16 

0.13 

0.21 

0.12 

0.19 

0.40 

0.38 

0.26 

0.28 

0.26 

0.20 

0.20 

0.15 

0.13 

0.20 

0.11 

0.18 

0.37 

0.35 

0.25 

0.26 

0.25 

0.19 

0.19 

0.15 

0.13 

0.19 

0.11 

0.17 

0.28 

0.27 

0.21 

0.21 

0.21 

0.17 

0.16 

0.13 

0.12 

0.17 

0.10 

0.15 

0.23 

0.22 

0.18 

0.18 

0.18 

0.15 

0.14 

0.12 

0.11 

0.15 

0.095 

0.14 

0.37 
0.36 

0.35 
0.34 

0.25 
0.24 

0.26 
0.25 

0.25 
0.24 

0.20 
0.19 

0.19 
0.19 

0.15 
0.15 

0.13 
0.13 

0.20 
0.19 

0.11 
0.11 

0.18 
0.17 

0.35 

0.33 

0.24 

0.25 

0.24 

0.19 

0.18 

0.14 

0.12 

0.19 

0.11 

0.17 

0.28 

0.26 

0.20 

0.21 

0.20 

0.17 

0.16 

0.13 

0.11 

0.17 

0.10 

0.15 

0.61 

0.56 

0.34 

0.36 

0.34 

0.25 

0.24 

0.18 

0.15 

0.25 

0.13 

0.22 

0.51 

0.47 

0.31 

0.32 

0.31 

0.23 

0.22 

0.17 

0.14 

0.23 

0.12 

0.20 

0.41 

0.38 

0.26 

0.28 

0.26 

0.20 

0.20 

0.15 

0.13 

0.21 

0.11 

0.18 

^Calculations  include  cement  mortar  (J^  in.)  between  veneer  or  facing  and  backing. 
•Based  on  one  air  cell  in  direction  of  heat  flow. 

/A  waterproof  membrane  should  be  provided  between  the  outer  material  and  the  insulation  fill  to 
prevent  possible  wetting  by  absorption  and  a  subsequent  lowering  of  efficiency. 

103 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

TABLE  5.    COEFFICIENTS  OF  TRANSMISSION  ( U)  OF 
VARIOUS  TYPES  OF  FRAME  CONSTRUCTION^ 


These  coefficients  are  expressed  in  Bin  per  hour  per  square  foot  per 
degree  Fahrenheit  difference  in  temperature  between  the  air  on  the  two 
sides,  and  are  based  on  a  wind  Telocity  of  16  mph. 


TYPICAL 
CONSTRUCTION 


EXTERIOR  FINISH 


TYPE  OF  SHEATHING 


voop 


1  in.  Wood* 


Wood  Siding  or  Clapboard 


in.  Rigid  Insulation 


in.  Plaster  Board 


W00.D 


1  in.  Wood* 


Wood  Shingles 


in.  Rigid  Insulation* 


in.  Plaster  Board* 


1  in.  Wood* 


Stucco 


in.  Rigid  Insulation 


/HEAWNQ- 


in.  Plaster  Board 


1  in.  Wood* 


Brick/  Veneer 


in.  Rigid  Insulation 


in.  Plaster  Board 


41 


42 


43 


44 


45 


47 


48 


49 


50 


51 


52 


^Computed  from  factors  marked  by  *  in  Table  2. 

6These  coefficients  may  alsoibe^used  with  sufficient  accuracy  for  plaster  on  wood  lath  or  plaster  on 
plaster  board. 

'Based  on  the  actual  width  of  2  by  4  studding,  namely,  3£i  in. 


104 


CHAPTER  5 — HEAT  TRANSMISSION  COEFFICIENTS  AND  TABLES 


INTERIOR  FINISH 


No  INSULATION  BETWEEN  STUDDING 


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0.27 

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0.21 

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0.43 

0.40 

0.26 

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0.28 

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0.24 

0.27 

0.28 

0.27 

0.20 

0.15 

0.12 

0.21 

0.21 

0.062 

0.18 

0.25 

0.26 

0.25 

0.19 

0.15 

0.11 

0.19 

0.20 

0.061 

0.18 

0.35 

0.37 

0.35 

0.24 

0.18 

0.13 

0.25 

0.25 

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INSULATION  BETWEEN  STUDDING 


<fY"eIIow'  pine  or  fir — actual  thickness  about  K/&  in. 
•Furring  strips  between  wood  shingles  and  sheathing. 

•''Small  air  space  and  mortar  between  building  paper  and  brick  veneer  neglected. 

*A  waterproof  membrane  should  Tbe  provided*  between  £he  outer  material  and  the  insulation  fill  to 
prevent  possible  wetting  by  absorption  and  a  subsequent  towering1  of  efficiency.  . 

I05: 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


TABLE  6.    COEFFICIENTS  OF  TRANSMISSION  (U)  OF  FRAME  INTERIOR  WALLS 

AND  PARTITIONS^ 

Coefficients  are  expressed  in  Btu  Per  hour  per  square  foot  per  degree  Fahrenheit  difference  in  temperature 
between  the  air  on  the  two  sides,  and  are  based  on  still  air  (no  'wind)  conditions  on  both  sides. 


TYPICAL 
^LA/T 

PU 

CONSTRUCTION 
ER,     yTUK/ 

feSu 

WALL 
No. 

SLVGLE 
PARTITION 
(FINISH 
ON  ONE 
SIDE  OP 
STUDDING) 

DOUBLE  PARTITION 
(FINISHED  ON  BOTH  SIDES  OP  STUDDING) 

Air 
Space 
Between 
Studding 

Flaked 
Gypsum 
Fill* 
Between 
Studding 

Rock 
Wool 
Fill* 
Between 
Studding 

l/z-in. 
Flexible 
Insulation 
Between 
Studding 
(One  Air 
Space) 

Stud  Space  Faced 
One     Side     with 
Bright  Aluminum 
Foil 

TYPE  OF  WALL 

A 

B 

C 

D 

E 

F 

Wood  Lath  and  Plaster 

On  Studding 

53 

0.62 

0.34 

0.11 

0.065 

0.21 

0.24 

Metal  Lath  and  Plaster* 

On  Studding 

54 

0.69 

0.39 

0.11 

0.066 

0.23 

0.26 

Plaster  Board  (%  in.)  and 
Plaster**  On  Studding 

55 

0.61 

0.34 

0.10 

0.065 

0.21 

0.24 

$4  in.  Rigid  Insulation  and 

Plaster*  On  Studding 

56 

0.35 

0.18 

0.083 

0.056 

0.14 

0.15 

1  in.  Rigid  Insulation  and 

Plaster*  On  Studding 

57 

0.23 

0.12 

0.066 

0.048 

0.097 

0.10 

IK  in-  Corkboard  and 

Plaster*  On  Studding 

58 

0.16 

0.081 

0.052 

0.040 

0.070 

0.073 

2  in.  Corkboard  and 
Plaster*  On  Studding 

59 

0.12 

0.063 

0.045 

0.035 

0.057 

0.059 

•Computed  from  factors  marked  by 
^Thickness  assumed  3jhf  in. 


*  in  Table  2.        'Plaster  on  metal  lath  assumed  %-in.  thick. 
^Plaster  assumed  K-in.  thick. 


TABLE  7.    COEFFICIENTS  OF  TRANSMISSION  (U)  OF  MASONRY  PARTITIONS* 

Coefficients  are  expressed  in  Btu  per  hour  Per  square  foot  Per  degree  Fahrenheit  difference  in  temperature 
between  the  air  on  the  two  sides,  and  are  based  on  still  air  (no  wind)  conditions  on  both  sides. 


TYPICAL  CONSTRUCTION 

1 

Ste^*! 

No. 

PLAIN  WALLS 
(No  PLASTER) 

WALLS 
PLASTERED 
ON  ONE  SIDE 

WALLS 
PLASTERED 
ON  BOTH  SIDES 

||   .u.'""- 

r/  :        -.jig-. 

TYPE  OF  WALL 

A 

B 

C 

4-in.  Hollow  Clay  Tile 

60 

0.45 

0.42 

0.40 

4-in.  Common 

Brick 

61 

0.50 

0.46 

0.43 

4-in.  Hollow  Gypsum  Tile 

62 

0.30 

0.28 

0.27 

2-in.  Solid  Plaster 

63 



0.53 

•Computed  from  factors  marked  by  *  in  Table  2. 

106 


CHAPTER  5  —  HEAT  TRANSMISSION  COEFFICIENTS  AND  TABLES 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


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oi      ^  **• 

z  ^ 

O     •«  g 

°1 

" 

t-0090 

OOit-I> 

iS 

(M      T^      CO       <N 

(M           (N           TH           TH 

O  CJ 

SI 

ss 

THOOO 

00 

OOOO 

O      fe-5 

O 

Sx 

fa   ^^e 

*"' 

G"^ 

O      o« 

s  s 

S    ^ 
»o     s  « 

'"""'     1 

1 

-«eo^. 

«« 

l>      00      C^      O 

IH*0 

•g 

§  - 

•8^ 

w     -1 

.S  c 

to    ^ 

N"          3.2 

CO      -^ 

1  " 

§  S  1|  : 

^CDQOO 
r-4 

*« 

•<±l       CO      "^      00 

^          «"g 

1         ^^ 

*         le 

?  a 

•S              '"  p, 
*                 >2 

>>           -S  ^ 

O     •$ 

•°            3g 

"  2  "i 

I 

T3                       "oTS 
qj                    G  ctJ 

H       ^ 

js»J                   --3-2 

§      I 

I  r  ysfr  gz  | 

1 

«     '           o"« 

u     2      .. 

e  5  /1»  sal 

J 

2        si 

'ABLE  10.  COEFF 
Coefficients  are 

TYPICAL  CONSTETJi 

COHCR.^^  /FLOi 

YPB  AND  THICKNESS  OP 

Rigid  Insulation6 
Rigid  Insulation6 
Corkboard" 
Corkboard* 

Computed  from  facto; 
ssumed  %  in.  thick, 
.ssumed  *%  in.  thick. 
Lssumed  1  in.  thick, 
'he  figures  for  Nos.  5 
concrete,  Usually  tl 

i(3'tj/Z_'^&>- 

H 

H     . 

c 
o 

c 

0 

.5    .5    .S.  .S 

V  $  tF  ^  v  % 

I 

z 

rH        r-4        OJ        £M 

S 

109 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


TABLE  11.    COEFFICIENTS  OF  TRANSMISSION  (U)  OF  VARIOUS  TYPES 
OF  FLAT  ROOFS  COVERED  WITH  BUILT-UP  ROOFING* 


TYPICAL  CONSTRUCTION 


1 
1 

i          WITH  METAL  LATH 
WITHOUT  CEILINGS                              AND 
PLASTEH  CEILINGS* 

TYPE  OF  ROOF  DECK 

THICKITOSS 

OP 

ROOF 
DECK. 
(INCHES) 

No. 

•t^sr                                  /£*3T 

XOOFlKCj        /Tut                          ROOriNrfj     /THE, 

^jL::  -..-.•:  •  -.  ^  •  3r]                 [?P  f.-:-  ••?-•--:,>  « 

Precast  Cement  Tile 

1H 

1 

^/WPP^T/^                     UfiTirj>ri'- 

CtlUKfi'^ 

irVULAUON/                    „  ^B1Ir'/oUTuw/ 
*OOMN<fc            /                  F.oonft<s>           / 

Concrete 

-2 

2 

f^'^H'^'il'ft                ^••:fejd 

Concrete 

4 

3 

coMCRtTC.^                        concntTE./    jiTj 

Concrete 

Q 

4 

^ElLlMd/ 

1N/ULA.TION/                        ,e,li'iULAri('"/ 
lOOHNffj               /                       ROOFlflffl              / 

Wood 
Wood 

1* 

1  LjTfc 

5 
^ 

557J  )  )T  J  V  7  j»  >V  /J                       Si  ry/yy  yyf  y./.* 

'«,«>'                       p|-»'  g( 

Wood 
Wood 

#* 

4> 

7 
8 

CLtlllHtf^. 

iNJULtftOtt/                                 Itl/ULAtlfln/ 
RflflRHCi             7                       T.«OPIM<^           ^ 

Gypsum  Fiber  Concrete6 
(2  in  )  on  Plaster  Board 

f  «f  r/y  n-  :-.  -  :  -V-  j                  ^^  "^  *  ""  ?:*:'"''>] 

(H  in.) 

2H 

9 

fLAJTCR.  50AfcP^                         PUA/TCR.   MAH.P* 

Gypsum  Fiber  Concrete6 
(3  in.)  on  Plaster  Board 

CttUWtf'^              ' 

(H  in.) 

3% 

10 

MOHruTi7      ^g^""( 

Flat  Metal  Roofs 

Coefficient    of    transmis- 
sion  of  bare  corrugated 

<^]jjf'™^        |'E""?iPr; 

Btu  per  hour  per  square 
foot  of  projected  area  per 



11 

dElLTHd^ 

ference    in    temperature, 
based  on  an  outside  wind 
velocity  of  15  mph. 

°Computed  from  factors  marked  by  *  in  Table  2. 

^Nominal  thicknesses  specified — actual  thicknesses  used  in  calculations. 

*Gypsum  fiber  concrete — 87K  per  cent  gypsum,  12>£  per  cent  wood  fiber. 


110 


CHAPTER  5 — HEAT  TRANSMISSION  COEFFICIENTS  AND  TABLES 


Coefficients  are  expressed  in  Btu  per  hour  per  square  foot  per  degree 
Fahrenheit  difference  in  temperature  bet-ween  the  air  on  the  two  sides, 
and  are  based  on  an  outside  wind  velocity  of  15  mph. 


WITHOUT  CEILING-UNDER  SIDE  OF 
HOOF  EXPOSED 


WITH  METAL  LATH  AND 
PLASTER,  CEILINGS* 


c 

3 

i 

c 

c 

S 

a 

a 

s 

.0 

JS 

i 

"S 

1 

i 

| 

s 

o 

1 

s 

a 
1 

1 

a 
1 

Is 

4f 

o 

s 

.s 

J2 

1 

-o 

'.5s 

Pi 

| 

-c 
"5c 

3 

| 

"2 
"Sb 

3 

1 

3 

'S 

5 

a 
t 

| 

6 

i 

o 

a 

1 

A 

B 

c 

D 

E 

F 

G 

H 

I 

T 

K 

L 

M 

N 

0 

P 

0.84 

0.37 

0.24 

0.18 

0.14 

0.22 

0.16 

0.13 

0.43 

0.26 

0.19 

0.15 

0.12 

0.18 

0.14 

0.11 

0.82 
0.72 
0.64 

0.37 
0.34 
0.33 

0.24 
0.23 
0.22 

0.17 
0.17 
0.16 

0.14 
0.13 
0.13 

0.22 
0.21 
0.21 

0.16 
0.16 
0.15 

0.13 
0.12 
0.12 

0.42 
0.40 
0.37 

0.26 
0.25 
0.24 

0.19 
0.18 
0.18 

0.15 
0.14 
0.14 

0.12 
0.12 
0.11 

0.18 
0.17 
0.17 

0.14 
0.13 
0.13 

0.11 
0.11 
0.11 

0.49 
0.37 
0.32 
0.23 

0.28 
0.24 
0.22 
0.17 

0.20 
0.18 
0.16 
0.14 

0.15 
0.14 
0.13 
0.11 

0.12 
0.11 
0.11 
0.096 

0.19 
0.17 
0.16 
0.13 

0.14 
0.13 
0.12 
0.11 

0.12 
0.11 
0.10 
0.091 

0.32 
0.26 
0.24 
0.18 

0.21 
0.19 
0.17 
0.14 

0.16 
0.15 
0.14 
0.12 

0.13 
0.12 
0.11 
0.10 

0.11 
0.10 
0.097 
0.087 

0.15 
0.14 
0.13 
0.11 

0.12 
0.11 
0.11 
0.096 

0.10 
0.095 
0.092 
0.082 

0.40 

0.25 

0.18 

0.14 

0.12 

0.17 

0.13 

0.11 

0.27 

0.19 

0.15 

0.12 

0.10 

0.14 

0.12 

0.097 

0.32 

0.22 

0.16 

0.13 

0.11 

0.15 

0.12 

0.10 

0.23 

0.17 

0.14 

0.11 

0.097 

0.13 

0.11 

0.091 

0.95 

0.39 

0.25 

0.18 

0.14 

0.23 

0.17 

0.13 

0.46 

0.27 

0.19 

0.15 

0.12 

0.18 

0.14 

0.11 

*These  coefficients  may  be  used  with  sufficient  accuracy  for  wood  lath  and  plastert  or  plaster  board  and 
plaster  ceilings.  It  is  assumed  that  there  is  an  air  space  between  the  under  side  of  the  roof  deck  and  the 
upper  side  of  the  ceiling. 


Ill 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


II 

8  1 

g  | 

£  Is 

to     •§  £ 
O       f»«c 


§  1 

2  "5>lJ 

2  ll 

S  ^'S 


l§ 
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E 


aJ 


s 


< 


(•m 
pnB  (Tit  z) 


(in 
(•mi)  uoi 


j  pus 
(•m  £f)  uoi^nsaj  piSrjj 


is  V«l  P°°A1 


(•m 
(-ut  %) 


(pasodxg  BJ 
Sxnj] 


O        N 


S 

i 


H 


s  ill 


2 
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-Sl 


3|«S 


lll-s 


*o'~f  ca'Soo  bfl'— 


112 


CHAPTER  5 — HEAT  TRANSMISSION  COEFFICIENTS  AND  TABLES 


TABLE  13.    COEFFICIENTS  OF  TRANSMISSION  (Z7)  OF  DOORS,  WINDOWS  AND  SKYLIGHTS 

Coefficients  are  based  on  a  wind  velocity  of  15  mph,  and  are  expressed  in  Btu  per  hour  per  square  foot  per 
degree  Fahrenheit  difference  in  temperature  between  the  air  inside  and  outside  of  the  door,  window  or  skylight 

A.   Windows  and  Skylights 


U 


Single 

Double.... 
Triple 


1.13*.* 

0.45' 

0.281* 


B.  Solid  Wood  Doors** 


NOMINAL 
THICKNESS 
INCHES 

ACTTTiL 

THICKNESS 
INCHES 

17 

1 

% 

0.69 

1H 

IHe 

0.59 

1H 

We 

0.52 

IX 

1H 

0.51 

2 

1% 

0.46 

2H 

2H 

0.38 

3 

2^i 

0.33 

•See  Heating,  Ventilating  and  Air  Conditioning,  by  Harding  and  Willard,  revised  edition,  1932. 

*Computed  using  C  =  1.15  for  wood;/i  =  1.65  and/0  =  6.0. 

*It  is  sufficiently  accurate  to  use  the  same  coefficient  of  transmission  for  doors  containing  thin  wood 
panels  as  that  of  single  panes  of  glass,  namely,  1.13  Btu  per  hour  per  square  foot  per  degree  difference 
between  inside  and  outside  air  temperatures, 

While  most  building  materials  have  surfaces  which  show  similar 
characteristics  as  far  as  the  transmission  of  heat  is  concerned,  it  is  a  well- 
known  fact  that  certain  surfaces  such  as  aluminum  bronze,  gold  bronze, 
aluminum  foil,  or  in  fact  any  metallic,  highly  polished  surface  presents  a 
greater  resistance  to  heat  transmission  than  the  surface  of  the  average 
building  material. 

The  greater  heat  resistance  of  such  metallic  surfaces  is  due  primarily  to 
their  higher  reflectivity  and  consequent  lower  emissivity  of  radiant  heat. 
The  use  of  multiple  layers  of  metallic  surfaces,  combined  with  air  spaces 
of  low  resistance,  provides  a  definite  insulating  effect.  Factors2  for  air 
spaces  bounded  by  aluminum  foil  are  given  in  Table  2. 

Coefficients  of  transmission  of  various  types  of  wall,  ceiling,  floor  and 
roof  construction  with  aluminum  insulation  can  be  readily  calculated. 
The  present  installation  practice  indicates  that  air  spaces  of  J^  in.  to 
1J^  in.  are  preferred  but  manufacturers'  recommendations  should  be 
closely  followed  in  the  application  of  aluminum  foil  insulation. 

The  majority  of  the  conductivities  and  conductances  of  the  building 
materials  and  insulations  given  in  Table  2  were  determined  by  the  hot- 
plate method  of  testing3.  Attention  is  called  to  the  fact  that  conductivi- 
ties per  inch  of  thickness  of  materials  or  insulations  do  not  afford  a  true 
basis  for  comparison,  although  they  are  frequently  used  for  that  purpose. 


^Insulating  Value  of  Bright  Metallic  Surfaces,  by  F.  B.  Rowley  (A.S.H.V.E.  Journal  Section,  Heating, 
Piping  and  Air  Conditioning,  June,  1934,  p.  263). 

Standard  Test  Code  for  Heat  Transmission  through  Walls  (A.S.H.V.E.  TRANSACTIONS,  Vol.  34,  1928). 
See  also  Chapter  40. 

113 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

Correct  comparisons  should  take  into  consideration  many  different 
factors,  including  conductivities  or  conductances,  thicknesses  installed 
and  manner  of  installation,  while  the  selection  of  an  insulation  should  also 
give  consideration  to  structural  qualities,  as  well  as  to  material  and 
application  costs.  Fire,  vermin,  and  rot  resistance  are  other  important 
factors  to  be  considered  when  comparing  materials.  At  present  there  is 
no  universally  recognized  method  of  rating  insulations.  Conductivities 
and  conductances  of  building  materials  and  insulations  are  useful  to  the 
heating  engineer  in  determining  over-all  coefficients  of  heat  transmission 
of  walls,  floors,  roofs  and  ceilings. 

Computed  Transmission  Coefficients 

Computed  heat  transmission  coefficients  of  many  common  types  of 
building  construction  are  given  in  Tables  3  to  13,  inclusive,  each  con- 
struction being  identified  by  a  serial  number.  For  example,  the  coefficient 
of  transmission  (U)  of  an  8-in.  brick  wall  and  }4  in.  of  plaster  is  0.46,  and 
the  number  assigned  to  a  wall  of  this  construction  is  1-B,  Table  3. 

Example  1.  Calculate  the  coefficient  of  transmission  (U)  of  an  8-in.  brick  wall  with 
14  in  of  piaster  applied  directly  to  the  interior  surface,  based  on  an  outside  wind  exposure 
of  15  mph.  It  is  assumed  that  the  outside  course  is  of  face  brick  having  a  conductivity 
of  9.20,  and  that  the  inside  course  is  of  common  brick  having  a  conductivity  of  5.0,  the 
thicknesses  each  being  4  in.  The  conductivity  of  the  plaster  is  assumed  to  be  3.3,  and  the 
inside  and  outside  surface  coefficients  are  assumed  to  average  1.65  and  6.00,  respectively, 
for  still  air  and  a  15  mph  wind  velocity. 

Solution,  k  (face  brick)  =  9.20;  x  =  4.0  in.;  k  (common  brick)  «  5.0;  x  «  4.0  in.; 
k  (plaster)  -  3.3;  x  =  H  in.;/i  =  1.65  ;/0  =  6.0.  Therefore, 


U 


"6X)  "*"  9.20   "*"  5.0   ""*"  3.3  T  1.65 

1 

"    0.167  +  0.435  4-  0.80  +  0.152  +  0.606 

**  0.46  Btu  per  hour  per  square  foot  per  degree  Fahrenheit  difference  in  tempera- 
ture between  the  air  on  the  two  sides. 

The  coefficients  in  the  tables  were  determined  by  calculations  similar 
to  those  shown  in  Example  1,  using  Fundamental  Formulae  2,  3,  4  and  5 
and  the  values  of  k  (or  Ca) ,  fi,  fo  and  a  indicated  in  Table  2  by  asterisks. 
In  computing  heat  transmission  coefficients  of  floors  laid  directly  on  the 
ground  (Table  10),  only  one  surface  coefficient  (fi)  is  used.  For  example, 
the  value  of  U  for  a  1-in.  yellow  pine  floor  (actual  thickness,  25/32  in.) 
placed  directly  on  6-in.  concrete  on  the  ground,  is  determined  as  follows: 

=  0.48  Btu  per  hour  per  square  foot  per  degree  difference 


0.781         6.0 


1.65        0.80         12.0 
in  temperature  between  the  ground  and  the  air  immediately  above  the  floor. 

The  thicknesses  upon  which  the  coefficients  in  Tables  3  to  13,  inclusive, 
are  based  are  as  follows  : 

Brick  veneer  ........................................................................................  4      }n- 

Plaster  and  metal  lath  .............  ',  ..........................................................     %  m- 

114 


CHAPTER  5  —  HEAT  TRANSMISSION  COEFFICIENTS  AND  TABLES 

Plaster    (on    wood  lath,    plasterboard,   rigid   insulation,   board 

form,  or  corkboard)  ...................  .  ....................................................     3^  in. 

Slate  (roofing)  ...............  „  ........  „  ...........................................................     }/%  in. 

Stucco  on  wire  mesh  reinforcing  ......................................................  1       in. 

Tar  and  gravel  or  slag-surfaced  built-up  roofing  _______  .....................     %  in. 

1-in.  lumber  (S  2-S)  ............................................................................  2^2  ^ 

IJi-in.  lumber  (S-2-S)  .............................................  Ijf6  in. 

2-in  lumber  (S-2-S)...  .....................................  .  ...................................  1%  in. 

2H-in.  lumber  (S-2-S)  ................................................................    2j|  in. 

3-in.  lumber  (S-2-S)  ............................................................................  2%  in. 

4-in.  lumber  (S-2-S)  ...............................  .  ..........  .  .......................  ...  .....  ..  3      in. 

Finish  flooring  (maple  or  oak)  ...........  . 


Solid  brick  walls  are  based  on  4-in.  face  brick  and  the  remainder 
common  brick.  Stucco  is  assumed  to  be  1-in.  thick  on  masonry  walls. 
Where  metal  lath  and  plaster  are  specified,  the  metal  lath  is  neglected. 

Rigid  insulation  refers  to  the  so-called  board  form  which  may  be  used 
structurally,  such  as  for  sheathing.  Flexible  insulation  refers  to  the 
blankets,  quilts  or  semi-rigid  types  of  insulation. 

Actual  thicknesses  of  lumber  are  used  in  the  computations  rather  than 
nominal  thicknesses.  The  computations  for  wood  shingle  roofs  applied 
over  wood  stripping  are  based  on  1  by  4  in.  wood  strips,  spaced  2  in.  apart. 
Since  no  reliable  figures  are  available  concerning  the  conductivity  of 
Spanish  and  French  clay  roofing  tile,  of  which  there  are  many  varieties, 
the  figures  for  such  types  of  roofs  were  taken  the  same  as  for  slate  roofs,  as 
it  is  probable  that  the  values  of  U  for  these  two  types  of  roofs  will 
compare  favorably. 

The  coefficients  of  transmission  of  the  pitched  roofs  in  Table  12  apply 
where  the  roof  is  over  a  heated  attic  or  top  floor  so  the  heat  passes  directly 
ihrough  the  roof  structure  including  whatever  finish  is  applied  to  the 
underside  of  the  roof  rafters. 

Combined  Coefficients  of  Transmission 

If  the  attic  is  unheated,  the  roof  structure  and  ceiling  of  the  top  floor 
must  both  be  taken  into  consideration,  and  the  combined  coefficient  of 
transmission  determined.  The  formula  for  calculating  the  combined 
coefficient  of  transmission  of  a  top-floor  ceiling,  unheated  attic  space,  and 
pitched  roof,  per  square  foot  of  roof  area,  is  as  follows: 

TJ  Ur  X   Z7Ce  (. 

U  =   ttX£/r+Z7ce  (6) 

where 

Z7r  =  coefficient  of  transmission  of  the  roof. 
Z7ce  =  coefficient  of  transmission  of  the  ceiling. 
n  —  the  ratio  of  the  area  of  the  roof  to  the  area  of  the  ceiling. 

In  using  this  formula,  a  correction  factor  must  be  applied.  As  the 
amount  of  heat  transferred  through  an  air  space  is  proportional  to  the 
difference  of  the  fourth  powers  of  the  absolute  temperatures  of  the  surfaces 
enclosing  the  air  space,  a  greater  amount  of  heat  is  absorbed  or  emitted 
by  radiation  by  the  surfaces  enclosing  an  unheated  attic  than  by  the 
surfaces  of  a  wall  or  ceiling  in  a  room  under  still-air  conditions,  where  the 
surrounding  objects  are  only  slightly  higher  in  temperature  than  the 
interior  surfaces  of  the  walls  and  ceiling.  For  example,  the  average 

115 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

coefficient  of  a  surface  in  still  air  is  1.65  Btu  per  hour  per  square  foot  per 
degree  Fahrenheit,  whereas  the  average  coefficient  of  an  air  space  in  an 
outside  wall  is  about  1.10  Btu  per  hour  per  square  foot  per  degree  Fahren- 
heit difference  between  the  two  surfaces,  at  a  mean  temperature  of  40  F. 
An  air  space  coefficient  of  1.10  is  equivalent  to  a  surface  coefficient 
of  2.20  for  each  of  the  two  surfaces  enclosing  the  air  space,  where  the 
over-all  transmission  is  computed  by  using  the  coefficients  of  the  two 
surfaces  enclosing  the  air  space  instead  of  the  coefficient  of  the  air  space 
itself.  Hence,  in  determining  the  values  of  Ur  and  Z7Ce  to  be  used  in  the 
formula,  the  coefficients  for  the  surfaces  of  the  roof  and  ceiling  enclosing 
the  attic  should  be  increased  to  allow  for  the  additional  amount  of  heat 
transferred  by  radiation,  and  a  coefficient  of  2.20  may  be  used  with 
sufficient  accuracy  for  each  of  these  surfaces,  although  in  very  precise 
work  a  correction  should  be  made  to  allow  for  the  fact  that  the  area  of  a 
pitched  roof  over  an  unheated  attic  is  greater  than  the  area  of  the  ceiling, 
and  hence,  the  amount  of  heat  absorbed  by  radiation  by  each  square  foot 
of  roof  surface  is  less  than  is  given  off  by  radiation  by  each  square  foot  of 
ceiling  surface. 

If  the  unheated  attic  space  between  the  roof  and  ceiling  has  no  dormers, 
windows  or  vertical  wall  surfaces,  the  combined  coefficients  may  be  used 
for  determining  the  heat  loss  through  the  roof  construction  between  the 
attic  and  top-floor  ceiling,  but  it  should  be  noted  that  these  coefficients 
should  be  multiplied  by  the  roof  area  and  not  by  the  ceiling  area.  If  the 
unheated  attic  contains  windows,  ventilators  or  vertical  wall  surfaces, 
which  would  tend  to  reduce  temperature  in  the  attic  to  a  temperature 
approaching  or  equaling  the  outside  temperature,  the  roof  should  be 
neglected  and  only  the  top-floor  ceiling  construction  and  the  correspond- 
ing ceiling  area  taken  into  consideration,  using  the  coefficients  given  in 
Tables  8  or  9.  Where  there  are  no  dormers,  doors,  or  windows,  and  when 
the  transmission  coefficients  of  the  roof  and  the  ceiling  are  approximately 
the  same,  the  value  of  the  attic  temperature  may  be  taken  as  an  average 
between  the  inside  and  the  outside  temperature. 

Basements  and  Unheated  Rooms 

The  heat  loss  through  floors  into  basements  and  into  unheated  rooms 
kept  closed  may  be  computed  by  assuming  a  temperature  for  these  rooms 
of  32  F. 

Additional  information  on  the  inside  and  outside  temperatures  to  be 
used  in  heat  loss  calculations  is  given  in  Chapter  7. 

REFERENCES 

A.S.H.V.E.  research  paper  entitled  Wind  Velocity  Gradients  Near  a  Surface  and  Their  Effect  on  Film 
Conductance,  by  F.  C.  Houghten  and  Paul  McDermott  (A.S.H.V.E.  TRANSACTIONS,  Vol.  37,  1931). 

A.S.H.V.E.  research  paper  entitled  Surface  Conductances  as  Affected  by  Air  Velocity,  Temperature  and 
Character  of  Surface,  by  F.  B.  Rowley,  A.  B.  Algren  and  J.  L.  Blackshaw  (A.S.H.V.E.  TRANSACTIONS, 
Vol.  36,  1930). 

A.S.H.V.E.  research  paper  entitled  Effects  of  Air  Velocities  on  Surface  Coefficients,  by  F.  B.  Rowley, 
A.  B.  Algren  and  J.  L.  Blackshaw  (A.S.H.V.E.  TRANSACTIONS,  Vol.  36,  1930). 

A.S.H.V.E.  research  paper  entitled  Conductivity  of  Concrete,  by  F.  C.  Houghten  and  Carl  Gutberlet 
(A.S.H.V.E.  TRANSACTIONS,  Vol.  37,  1931). 

A.S.H.V.E.  research  paper  entitled  Surface  Coefficients  as  Affected  by  Direction  of  Wind,  by  F.  B. 
Rowley  and  W.  A.  Eckley  (A.S.H.V.E.  TRANSACTIONS,  Vol.  37,  1931). 

A.S.H.V.E.  research  paper  entitled  Thermal  Resistance  of  Air  Spaces,  by  F.  B.  Rowley  and  A.  B.  Algren 
(A.S.H.V.E.  TRANSACTIONS,  Vol.  35,  1929). 

116 


CHAPTER  5 — HEAT  TRANSMISSION  COEFFICIENTS  AND  TABLES 

A.S.H.V.E.  research  paper  entitled  The  Heat  Conductivity  of  Wood  at  Climatic  Temperature  Dif- 
ferences, by  F.  B.  Rowley  (A.S.H.V.E.  Journal  Section,  Heating,  Piping  and  Air  Conditioning,  June,  1933). 

A.S.H.V.E.  research  paper  entitled  Insulating  Value  of  Bright  Metallic  Surfaces,  by  F.  B.  Rowley 
(A.S.H.V.E.  Journal  Section,  Heating,  Piping  and  Air  Conditioning,  June,  1934). 

Heat  Transmission  through  Building  Materials,  by  F.  B.  Rowley  and  A.  B.  Algren,  University  of  Min- 
nesota Engineering  Experiment  Station  Bulletin  No.  8. 

Insulating  Effect  of  Successive  Air  Spaces  Bounded  by  Bright  Metallic  Surfaces,  by  L.  W.  Schad  (A.S.H.- 
V.E. TRANSACTIONS,  Vol.  37,  1931). 

Importance  of  Radiation  in  Heat  Transfer  through  Air  Spaces,  by  E.  R.  Queer  (A.S.H.V.E.  TRANS- 
ACTIONS, Vol.  38,  1932). 

Properties  of  Metal  Foil  as  an  Insulating  Material,  by  J.  L.  Gregg  (Refrigerating  Engineering,  May,  1932). 

Thermal  Insulation  with  Aluminum  Foil,  by  R.  B.  Mason  (Industrial  and  Engineering  Chemistry, 
March,  1933). 

Heating,  Ventilating  and  Air  Conditioning,  by  Harding  and  Willard,  Revised  Edition,  1932. 

Thermal  Insulation  of  Buildings,  Technical  Paper  No.  11  (American  Architect,  May,  1934). 

House  Insulation,  Its  Economies  and  Application,  by  Russell  E.  Backstrom  (Report  of  the  National 
Committee  on  Wood  Utilization,  United  States  Government  Printing  Office,  1931). 

Heat  Insulation  as  Applied  to  Buildings  and  Structures,  by  E.  A.  Allcut,  University  of  Toronto,  1934. 


PROBLEMS  IN  PRACTICE 

1  •  What  is  the  conductance  of  a  1-in.  air  space,  faced  with  common  building 
materials,  at  a  mean  temperature  of  50  F? 

1.152  (Table  1). 

2  •  What  is  the  conductivity  of  face  brick? 

9.20  (Table  2). 

3  •  What  is  the  conductance  of  wood  shingles? 

1.28  (Table  2). 

4  •  What  is  the  over-all  coefficient  of  transmission  U  for  a  solid  brick  wall 
12 -in.  thick  with  plaster  on  wood  lath,  furred? 

0.24  (Table  3,  Wall  2C). 

5  •  Find  the  value  of  U  for  a  6-in.  concrete  wall  with  plaster  on  metal  lath 
attached  to  2 -in.  furring  strips  with  flanged  J-^-in.  blanket  insulation. 

0.23  (Table  3,  Wall  12L). 

6  •  Find  the  value  of  U  for  a  wood  siding  wall  with  an  interior  finish  of  J£-in. 
plaster  on  metal  lath;  sheathing  thickness,  2%%  in. 

0.26  (Table  5,  Wall  41B). 

7  •  What  value  of  U  should  be  used  for  a  brick  veneer  wall  with  H-in.  rigid 
insulation  sheathing  finished  on  the  interior  with  plaster  on  J^-in.  rigid  insu- 
lation? 

0.19  (Table  5,  Wall  51D). 

8  •  What  value  of  U  should  be  used  in  computing  the  heat  loss  from  an  attic 
through  a  floor  of  yellow  pine  on  joists  with  a  ceiling  of  metal  lath  and  plaster? 

0.30  (TableS,  Floor  2B). 

9  •  What  is  the  over-all  heat  transfer  coefficient  for  a  6-in.  concrete  floor  with 
no  insulation  and  with  yellow  pine  flooring  on  sleepers  resting  on  concrete? 

0.33  (Table  10,  Floor  2B). 

10  •  What  is  the  coefficient  U  for  a  flat  roof  of  4-in.  concrete  with  a  metal  lath 
and  plaster  ceiling  insulated  with  1-in.  cork  board? 

0.17  (Table  11,  Roof  3N). 

117 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

11  •  A  solid  12-in.  common  brick  wall  is  finished  on  the  inside  with  J^-in. 
insulation  plaster  base,  and  J^-in.  plaster;  the  plaster  base  is  furred  1  in.  from 
the  brick;  k  for  insulating  material  =  0.34.     Calculate  the  over-all  coefficient 
V. 

fi  =  1.65;  /0  =  6.00;  k  for  brick  =  5.00;  a  for  1-in.  air  space  =  1.1 

Over-all  heat  resistance  =  R  =  -^  +  ^  +  rT  +  ~lf  +  'ir==  5-703 

l.OO  O.O  1.1  O  O 

u  =  -4- 

K 

12  •  A  wall  is  built  with  two  layers  of  >£-in.  insulating  material  spaced  1  in. 
apart;  the  air  space  is  lined  on  one  side  with  bright  aluminum  foil;  mean 
temperature  is  40  F;  still  ah*  on  both  sides  of  wall;  k  for  insulating  material 
is  0.34.    Calculate  the  value  of  U. 


fi  =  1.65; /0  =  1.65;  a  «  0.46 
1       ,     0.5     ,       1       ,     0.5     , 


-  6.327 


~   1.65   ^  0.34   7  0.46  ^  0.34     '     1.65 
U  =  4-  =  °-158 

JK. 

13  •  What  is  the  inside  surface  temperature  of  a  6-in.  solid  concrete  wall? 
Inside  air,  70  F;  outside  ah*,  —20  F  with  15  mph  wind. 

The  temperature  drop  from  point  to  point  through  a  wall  is  directly  proportional  to  the 
heat  resistance. 

fi  =  1.65;  k  for  concrete  =  12; /0  =  6.0 

i  ft  i 

Over-all  resistance  R  =    1-5F  +  T7  +  ^7i  =  L27 

i.OO  JLj£  O.U 

Temperature  drop,  inside  air  to  surface  _    1.65 
Temperature  drop,  air  to  air  1.27 

90 

Temperature  drop,  inside  air  to  surface  =  --.-^-rr  ;,  />r    =  43   i 

JL.Z/    ^\  I.uo 

70  —  43  =  27  F,  inside  surface  temperature  of  wall. 

14  •  How  many  inches  of  insulating  material  having  a  conductivity  of  0.30 
would  be  required,  for  the  wall  of  Question  3,  to  raise  the  inside  surface  tem- 
perature to  60  F? 

Temperature  drop,  air  to  inside  surface  =  10  F;  temperature  drop,  inside  surface  to  out- 
side air  =  80  F.    Therefore,  the  heat  resistance  from  inside  wall  surface  to  outside  air 

must  be  eight  times  that  from  inside  air  to  inside  wall  surface,  or  8  X   ..  „*  =  4.85.   The 
resistance  for  added  material  is,  therefore, 


-    w  +         -  4'19 

4.19  X  0.30  =  1.25  in.  of  insulation. 


118 


Chapter  6 

AIR  LEAKAGE 

Nature  of  Air  Infiltration,  Air  Leakage  Through  Walls,  Window 
Leakage,  Wind  Velocity  to  be  Selected,  Crack  used  for  Computa- 
tions, Multi-Story  Buildings,  Heat  Equivalent  of  Air  Entering 
by  Infiltration 

AIR  leakage  losses  are  those  resulting  from  the  displacement  of  heated 
air  in  a  building  by  unheated  outside  air,  the  interchange  taking 
place  through  various  apertures  in  the  building,  such  as  cracks  around 
doors  and  windows,  fireplaces  and  chimneys.    This  leakage  of  air  must  be 
considered  in  heating  and  cooling  calculations.     (See  Chapters  7  and  8.) 

THE  NATURE  OF  AIR  FILTRATION 

The  natural  movement  of  air  through  building  construction  is  due  to 
two  causes.  One  is  the  pressure  exerted  by  the  wind;  the  other  is  the 
difference  in  density  of  outside  and  inside  air  because  of  differences  in 
temperature. 

The  wind  causes  a  pressure  to  be  exerted  on  one  or  two  sides  of  a 
building.  As  a  result,  air  comes  into  the  building  on  the  windward  side 
through  cracks  or  porous  construction,  and  a  similar  quantity  of  air 
leaves  on  the  leeward  side  through  like  openings.  In  general  the  resis- 
tance to  air  movement  is  similar  on  the  windward  to  that  on  the  leeward 
side.  This  causes  a  building  up  of  pressure  within  the  building  and  a 
lesser  air  leakage  than  that  experienced  in  single  wall  tests  as  determined 
in  the  laboratory.  It  is  assumed  that  actual  building  leakages  owing  to 
this  building  up  of  pressure  will  be  80  per  cent  of  laboratory  test  values. 
While  there  are  cases  where  this  is  not  true,  tests  in  actual  buildings 
substantiate  the  factor  for  the  general  case.  Tests  on  mechanically 
ventilated  classrooms  of  average  construction  have  shown  that  air 
infiltration  acts  quite  independently  of  the  planned  air  supply.  Accor- 
dingly, the  heating  or  cooling  load  owing  to  air  infiltration  from  natural 
causes  should  be  considered  in  addition  to  the  ventilating  load. 

The  air  exchange  owing  to  temperature  difference,  inside  to  outside,  is 
not  appreciable  in  low  buildings.  In  tall,  single  story  buildings  with 
openings  near  the  ground  level  and  near  the  ceiling,  this  loss  must  be 
considered.  Also  in  multi-storied  buildings  it.  is  a  large  item  unless  the 
sealing  between  various  floors  and  rooms  is  quite  perfect.  This  tempera- 
ture effect  is  a  chimney  action,  causing  air  to  enter  through  openings  at 
lower  levels  and  to  leave  at  higher  levels. 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

A  complete  study  of  all  of  the  factors  involved  in  air  movement  through 
building  constructions  would  be  very  complex.  Some  of  the  complicating 
factors  are:  the  variations  in  wind  velocity  and  direction;  the  exposure  of 
the  building  with  respect  to  air  leakage  openings  and  with  respect  to 
adjoining  buildings;  the  variations  in  outside  temperatures  as  influencing 
the  chimney  effect;  the  relative  area  and  resistance  of  openings  on  the 
windward  and  leeward  sides  and  on  the  lower  floors  and  on  the  upper 
floors;  the  influence  of  a  planned  air  supply  and  the  related  outlet  vents; 
and  the  variation  from  the  average  of  individual  building  units.  A  study 
of  infiltration  points  to  the  need  for  care  in  the  obtaining  of  good  building 
construction,  or  unnecessarily  large  heat  losses  will  result. 


AIR  LEAKAGE  THROUGH  WALLS 

Table  I1  gives  data  on  infiltration  through  brick  and  frame  walls.  The 
brick  walls  listed  in  this  table  are  walls  which  show  poor  workmanship 
and  which  are  constructed  of  porous  brick  and  lime  mortar.  For  good 
workmanship,  the  leakage  through  hard  brick  walls  with  cement-lime 
mortar  does  not  exceed  one- third  the  values  given.  These  tests  indicate 
that  plastering  reduces  the  leakage  by  about  96  per  cent;  a  heavy  coat  of 
cold  water  paint,  50  per  cent;  and  3  coats  of  oil  paint  carefully  applied, 
28  per  cent.  The  infiltration  through  walls  ranges  from  6  to  25  per  cent 
of  that  through  windows  and  doors  in  a  10-story  office  building,  with 
imperfect  sealing  of  plaster  at  the  baseboards  of  the  rooms.  With  perfect 
sealing  the  range  is  from  0.5  to  2.7  per  cent  or  a  practically  negligible 
quantity,  which  indicates  the  importance  of  good  workmanship  in  proper 
sealing  at  the  baseboard.  It  will  be  noted  from  Table  1,  that  the  in- 
filtration through  properly  plastered  walls  can  be  neglected. 

TABLE  1.  INFILTRATION  THROUGH  WALLS 

Expressed  in  cubic  feet  per  square  foot  per  hour* 


WIND  VBLOCTTT,  MILES  PER  Hora 


TYPE  OP  WALL 

5 

10 

is 

20 

25 

30 

8«  m.  Brick  WalL_{g£—L: 

1.75 

0.017 

4.20 
0.037 

7.85 
0.066 

12.2 
0.107 

18.6 
0.161 

22.9 
0.236 

13  in.  Brick  Wall  {$£^£1 

1.44 
0.005 

3.92 
0.013 

7.48 
0.025 

11.6 
0.043 

16.3 
0.067 

21.2 
0.097 

Frame  Wall,  with  lath  and  plasterb 

0.03 

0.07 

0.13 

0.18 

0.23 

0.26 

aThe  values  in  this  table  are  20  per  cent  less  than  test  values  to  allow  for  building  up  of  pressure  in  rooms 
and  are  based  on  test  data  reported  in  A.S.H.V.E.  research  papers  entitled  Air  Infiltration  Through  Various 
Types  of  Brick  Wall  Construction,  and  Air  Infiltration  Through  Various  Types  of  Wood  Frame  Con- 

struction.    (See  References  on  pages  128  and  129). 

bWall  construction:  Bevel  siding  painted  or  cedar  shingles,  sheathing,  building  paper,  wood  lath  and 
3  coats  gypsum  plaster. 


*Air  Infiltration  through  Various  Types  of  Brick  Wall  Construction,  by  Larson,  Nelson  and  Braatz 
A.S.H.V.E.  TRANSACTIONS,  Vol.  36r  1930). 

120 


CHAPTER  6 — AIR  LEAKAGE 


0  20          40  60  SO          /OO         /&         i4O         /6O         /SO        ZOO        22O        24O        ^ffO        2BO        3OO 

INFILTRATION  w  C  FH.  PER  SQ.  FT  OF  WALL 

FIG.  1.   INFILTRATION  THROUGH  VARIOUS  TYPES  OF  SHINGLE  CONSTRUCTION 

The  value  of  building  paper  when  applied  between  sheathing  and 
shingles  is  indicated  by  Fig.  1,  which  represents  the  effect  on  outside 
construction  only,  without  lath  and  plaster.  The  effectiveness  of  plaster 
properly  applied  is  no  justification  for  the  use  of  low  grade  building  paper 
or  of  the  poor  construction  of  the  wall  containing  it.  Not  only  is  it 
difficult  to  secure  and  maintain  the  full  effectiveness  of  the  plaster  but 
also  it  is  highly  desirable  to  have  two  points  of  high  resistance  to  air  flow 
with  an  air  space  between  them. 


^0.05 


/OO          /£0          MO        S&>         /8O 
M9    C.f.M  P£f9  5<>,  FT.    Of   WALL 


FIG.  2.  INFILTRATION  THROUGH  SINGLE  SURFACE  WALLS  USED  IN  FARM  AND 
OTHER  SHELTER  BUILDINGS 

121 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

The  amount  of  infiltration  that  may  be  expected  through  simple  walls 
used  in  farm  and  other  shelter  buildings,  is  shown  in  Fig.  2.  The  infil- 
tration there  indicated  is  that  determined  in  the  laboratory  and  should  be 
multiplied  by  the  factor  0.80  to  give  proper  working  values. 

WINDOW  LEAKAGE 

The  amount  of  infiltration  for  various  types  of  windows  is  given  in 
Table  2.  The  fit  of  double-hung  wood  windows  is  determined  by  crack 
and  clearance  as  illustrated  in  Fig.  3.  The  length  of  the  perimeter  opening 
or  crack  for  a  double-hung  window  is  equal  to  three  times  the  width  plus 
two  times  the  height,  or  in  other  words,  it  is  the  outer  sash  perimeter 
length  plus  the  meeting  rail  length.  Values  of  leakage  shown  in  Table  2 
for  the  average  double-hung  wood  window  were  determined  by  setting 
the  average  measured  crack  and  clearance  found  in  a  field  survey  of  a 
large  number  of  windows  on  nine  windows  tested  in  the  laboratory.  In 
addition,  the  table  gives  figures  for  a  poorly  fitted  window.  All  of  the 
figures  for  double-hung  wood  windows  are  for  the  unlocked  condition. 
Just  how  a  window  is  closed,  or  fits  when  it  is  closed,  has  considerable 
influence  on  the  leakage.  The  leakage  will  be  high  if  the  sash  are  short, 
if  the  meeting  rail  members  are  warped,  or  if  .the  frame  and  sash  are  not 
fitted  squarely  to  each  other.  It  is  possible  to  have  a  window  with 
approximately  the  average  crack  and  clearance  that  will  have  a  leakage 
at  least  double  that  of  the  figures  shown.  Values  for  the  average  double- 
hung  wood  window  in  Table  2  are  considered  to  be  easily  obtainable 
figures  provided  the  workmanship  on  the  window  is  good.  Should  it  be 
known  that  the  windows  under  consideration  are  poorly  fitted,  the  larger 
leakage  values  should  be  used.  Locking  a  window  generally  decreases  its 
leakage,  but  in  some  cases  may  push  the  meeting  rail  members  apart  and 
increase  the  leakage.  On  windows  with  large  clearances,  locking  will 
usually  reduce  the  leakage. 

Wood  casement  windows  may  be  assumed  to  have  the  same  unit 
leakage  as  for  the  average  double-hung  wood  window  when  properly 
fitted.  Locking,  a  normal  operation  in  the  closing  of  this  type  of  window, 
maintains  the  crack  at  a  low  value. 

For  metal  pivoted  sash,  the  length  of  crack  is  the  total  perimeter  of  the 
movable  or  ventilating  sections.  Frame  leakage  on  steel  windows  may  be 
neglected  when  they  are  properly  grouted  with  cement  mortar  into  brick 
work  or  concrete.  When  they  are  not  properly  sealed,  the  linear  feet  of 
sash  section  in  contact  with  steel  work  at  mullions  should  be  figured  at 
25  per  cent  of  the  values  for  industrial  pivoted  windows  as  given  in 
Table  2. 

Leakage  values  for  storm  sash  are  given  in  Figs.  4  and  5.  When  storm 
sash  are  applied  to  well  fitted  windows,  very  little  reduction  in  infiltration 
is  secured,  but  the  application  of  the  sash  does  give  an  air  space  which 
reduces  the  heat  transmission  and  helps  prevent  the  frosting  of  the 
windows.  When  storm  sash  are  applied  to  poorly  fitted  windows,  a 
reduction  in  leakage  of  50  per  cent  may  be  secured. 

Doors  vary  greatly  in  fit  because  of  their  large  size  and  tendency  to 
warp.  For  a  well  fitted  door,  the  leakage  values  for  a  poorly  fitted  double- 
hung  wood  window  may  be  used.  If  poorly  fitted,  twice  this  figure  should 

122 


CHAPTER  6 — AIR  LEAKAGE 


TABLE  2.    INFILTRATION  THROUGH  WINDOWS 
Expressed  in  Cubic  Feet  per  Foot  of  Crack  per  Hour3- 


TYPE  OF  WINDOW 

REMARKS 

WIND  VELOCITY,  MILES  PER  HOTTE 

5 

10 

15 

20 

25 

30 

Double-Hung 
Wood  Sash 
Windows 
(Unlocked) 

Around  frame  in  masonry  wall  — 
not  calkedb  

3.3 

8.2 

14.0 

20.2 

27.2 

34.6 

Around  frame  in  masonry  wall  — 
calkedb  

0.5 

1.5 

2.6 

3.8 

4.8 

5.8 

Around  frame   in   wood   frame 
construction  D  

2.2 

6.2 

10.8 

16.6 

23.0 

30.3 

Total  for  average  window,  non- 
weatherstripped,  Me-in.  crack 
and    %4-in.    clearance0.     In- 
cludes wood  frame  leakaged  

6.6 

21.4 

39.3 

59.3 

80.0 

103.7 

Ditto,  weatherstrippedd»  

4.3 

15.5 

23.6 

35.5 

48.6 

63.4 

Total  for  poorly  fitted  window, 
non-weatherstripped,    %  2  -in  . 
crack  and  %2-in.  clearance6. 
Includes  wood  frame  leakaged. 

26.9 

69.0 

110.5 

153.9 

199.2 

249.4 

Ditto,  weatherstrippedd  

5.9 

18.9 

34.1 

51.4 

70.5 

91.5 

Double-Hung 
Metal 
WTindowsf 

Non-weatherstripped,  locked  
Non-weatherstrippedr  unlocked.. 
Weatherstripped,  unlocked  

20 
20 
6 

45 
47 
19 

70 
74 
32 

96 
104 
46 

125 
137 
60 

154 
170 
76 

Rolled 
Section 
Steel  Sash 
Windowsk 

Industrial  pi  voted,  s  He-in.  crack 
Architectural  projected,11  JNU-in. 
crack. 

52 
20 
14 

8 

108 
52 
32 

24 

176 
88 
52 
38 

244 
116 
76 
54 

304 
152 
100 
72 

372 
208 
128 
96 

Residential    casement,1     J^2-in. 
crack. 

Heavy  casement   section,    pro- 
jected, J  3^2"in.  crack.  

Hollow  Metal,  vertically  pivoted  window*.  

30 

88 

145 

186 

221 

242 

"The  values  given  in  this  table  are  20  per  cent  less  than  test  values  to  allow  for  building  up  of  pressure  in 
rooms,  and  are  based  on  test  data  reported  in  the  papers  listed  at  the  end  of  this  chapter. 

bThe  values  given  for  frame  leakage  are  per  foot  of  sash  perimeter  as  determined  for  double-hung  wood 
windows.  Some  of  the  frame  leakage  in  masonry  walls  originates  in  the  brick  wall  itself  and  cannot  be 
prevented  by  calking.  For  the  additional  reason  that  calking  is  not  done  perfectly  and  deteriorates  with 
time,  it  is  considered  advisable  to  choose  the  masonry  frame  leakage  values  for  calked  frames  as  the  average 
determined  by  the  calked  and  not-calked  tests. 

cThe  fit  of  the  average  double-hung  wood  window  was  determined  as  }£-m.  crack  and  %-in.  clearance  by 
measurements  on  approximately  600  windows  under  heating  season  conditions. 

dThe  values  given  are  the  totals  for  the  window  opening  per  foot  of  sash  perimeter  and  include  frame 
leakage  and  so-called  elsewhere  leakage.  The  frame  leakage  values  included  are  for  wood  frame  construction 
but  apply  as  well  to  masonry  construction  assuming  a  60  per  cent  efficiency  of  frame  calking. 

*A  J6-in.  crack  and  clearance  represents  a  poorly  fitted  window,  much  poorer  than  average. 

^Windows  tested  in  place  in  building. 

^Industrial  pivoted  window  generally  used  in  industrial  buildings.  Ventilators  horizontally  pivoted 
at  center  or  slightly  above,  lower  part  swinging  out. 

^Architectural  projected  made  of  same  sections  as  industrial  pivoted  except  that  outside  framing  member 
is  heavier,  and  refinements  in  weathering  and  hardware.  Used  in  semi- monumental  buildings  such  as  schools. 
Ventilators  awing  in  or  out  and  are  balanced  on  side  arms. 

[Of  same  design  and  section  shapes  as  sc-called  heavy  section  casement  but  of  lighter  weight. 

iMade  of  heavy  sections.    Ventilators  swing  in  or  out  and  stay  set  at  any  degree  of  opening. 

kWith  reasonable  care  in  installation,  leakage  at  contacts  where  windows  are  attached  to  steel  frame- 
work and  at  mulKons  is  negligible.  With  ?6-in.  crack,  representing  poor  installation,  leakage  at  contact 
with  steel  framework  is  about  one-third,  and  at  mullions  about  one-sixth  of  that  given  for  industrial  pivoted 
windows  in  the  table. 

123 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

be  used.  If  weathers  tripped,  the  values  may  be  reduced  one-half.  A 
single  door  which  is  frequently  opened,  such  as  might  be  found  in  a  store, 
should  have  a  value  applied  which  is  three  times  that  for  a  well  fitted 
door.  This  extra  allowance  is  for  opening  and  closing  losses  and  is  kept 
from  being  greater  by  the  fact  that  doors  are  not  used  as  much  in  the 
coldest  and  windiest  weather. 

CHOOSING  WIND  VELOCITY 

Although  all  authorities  do  not  agree  upon  the  value  of  the  wind  veloc- 
ity that  should  be  chosen  for  any  given  locality,  it  is  common  engineering 
practice  to  use  the  average  wind  velocity  during  the  three  coldest  months 
of  the  year.  Until  this  point  is  definitely  established  the  practice  of 
using  average  values  will  be  followed.  Average  wind  velocities  for  the 
months  of  December,  January  and  February  for  various  cities  in  the 
United  States  and  Canada  are  given  in  Table  2,  Chapter  7. 


FIG.  3.    DIAGRAM  ILLUSTRATING  CRACK  AND  CLEARANCE 

In  considering  both  the  transmission  and  infiltration  losses,  the  more 
exact  procedure  would  be  to  select  the  outside  temperature  and  the  wind 
velocity  corresponding  thereto,  based  on  Weather  Bureau  records,  which 
would  result  in  the  maximum  heat  demand.  Since  the  proportion  of 
transmission  and  infiltration  losses  varies  with  the  construction  and  is 
different  for  every  building,  the  proper  combination  of  temperature  and 
wind  velocity  to  be  selected  would  be  different  for  every  type  of  building, 
even  in  the  same  locality.  Furthermore,  such  a  procedure  would  neces- 
sitate a  laborious  cut-and-try  process  in  every  case  in  order  to  determine 
the  worst  combination  of  conditions  for  the  building  under  consideration. 
It  would  also  be  necessary  to  consider  heat  lag  due  to  heat  capacity  in  the 
case  of  heavy  masonry  walls,  and  other  factors,  to  arrive  at  the  most 
accurate  solution  of  the  problem.  Although  heat  capacity  should  be  con- 
sidered wherever  possible,  it  is  seldom  possible  to  accurately  determine  the 
worst  combination  of  outside  temperature  and  wind  velocity  for  a  given 
building  and  locality.  The  usual  procedure,  as  already  explained,  is  to 
select  an  outside  temperature  based  on  the  lowest  on  record  and  the 
average  wind  velocity  during  the  months  of  December,  January  and 
February. 

The  direction  of  prevailing  winds  may  usually  be  included  within  an 
angle  of  about  90  deg.  The  windows  that  are  to  be  figured  for  prevailing 

124 


CHAPTER  6 — AIR  LEAKAGE 


12 

a  u 

•?  ad 
1Q7 

<u 

§02 

< 

50.03 
45.63 
40.83 
35.4Q 
ZdSO 

, 
J 

C 

f 

d 

JA 

1 

d 

j 

\ 

1 

I 

/ 

1 

1 

/ 

I 

1 

1 

f 

i 

1 

1 

1 

1 

/ 

1 

J 

1 

1 

j 

1 

/ 

j 

I 

It 

A  -WITHOUT  STORM  SASH 
8*  STORM  SASH-  SUSPENDED 

C-  STORM  SASH-fASTEMEQ 

WITH  FOUR  TURN  BUTTONS 
D-  SAME  As  C  WITH  WOOL 
WEATHEZ-STR/P  APPLIED 
To  STORM  SASH 

y 

I 

/ 

I 

1 

f 

/ 

I 

II 

- 

1  > 

/ 

1 

— 

j 

I 

/ 

// 

/ 

ty 

0 

1 

$ 

1 

jp 

->         50         KX>        ISO        200      Z5Q      300 
INFILTRATION  CJ:H.Pex  roar  GrOwcx 

FIG.  4. 


INFILTRATION  THROUGH  SASH  PERIMETER  OF  WINDOW  WITH  AND  WITHOUT 
STORM  SASH  —  J^4-iN.  CRACK  AND  jH?2-iN.  CLEARANCE 


and  non-prevailing  winds  will  ordinarily  each  occupy  about  one-half  the 
perimeter  of  the  structure,  the  proportion  varying  to  a  considerable  extent 
with  the  plan  of  the  structure.  (See  discussion  of  wind  movement  in 
Chapter  4.) 


LZ 


07 


(26 


*Q4 


50 


A*W/THOUT  STORM  SASH 
5-  STORM  SASH  •  SUSPENDED 
C  *  'STORM  SASH  •  FASTENED 
Wrrn  Foue  TURN  BUTTONS 


JOO        150       200      Z50       500      350 
INFIUTZATION  Cf.H.  PER  FOOT  OF  CRACK 


400 


50.03 
45.65 


2430 


i 


FIG.  5.    INFILTRATION  THROUGH  SASH  PERIMETER  OF  WINDOW  WITH  AND  WITHOUT 
STORM  SASH— K-*N<  CRACK  AND  J^-IN.  CLEARANCE 


125 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

CRACK  USED  FOR  COMPUTATIONS 

In  no  case  should  the  amount  of  crack  used  for  computation  be  less 
than  half  of  the  total  crack  in  the  outside  walls  of  the  room.  Thus,  in  a 
room  with  one  exposed  wall,  take  all  the  crack;  with  two  exposed  walls, 
take  the  wall  having  the  most  crack ;  and  with  three  or  four  exposed  walls, 
take  the  wall  having  the  most  crack ;  but  in  no  case  take  less  than  half  the 
total  crack.  For  a  building  having  no  partitions,  whatever  wind  enters 
through  the  cracks  on  the  windward  side  must  leave  through  the  cracks 
on  the  leeward  side.  Therefore,  take  one-half  the  total  crack  for  com- 
puting each  side  and  end  of  the  building. 

The  amount  of  air  leakage  is  sometimes  roughly  estimated  by  assuming 
a  certain  number  of  air  changes  per  hour  for  each  room,  the  number  of 
changes  assumed  being  dependent  upon  the  type,  use  and  location  of  the 
room,  as  indicated  in  Table  3.  This  method  may  be  used  to  advantage  as 
a  check  on  the  calculations  made  in  the  more  exact  manner. 


TABLE  3.      AIR  CHANGES  TAKING  PLACE  UNDER  AVERAGE  CONDITIONS  EXCLUSIVE 
OF  AIR  PROVIDED  FOR  VENTILATION 


KIND  OF  BOOM  OB  BUILDING 

NUMBER  OP  Am  CHANGES 
TAKING  PLACE 
PER  HOUR 

Rooms,  1  side  exposed      

1 

Rooms,  2  sides  exposed 

\y> 

Rooms,  3  sides  exposed         

2 

Rooms  4  sides  exposed 

2 

Rooms  with  no  windows  or  outside  doors    

lAto  % 

Entrance  Halls 

2  to  3 

Reception  Halls                

2 

Living  Rooms  

1  to  2 

Dining  Rooms    .        

1  to  2 

Bath  Rooms 

2 

Drug  Stores           

2  to  3 

Clothing  Stores 

1 

Churches,  Factories,  Lofts,  etc. 

%  to  3 

MULTI-STORY  BUILDINGS 

In  tall  buildings,  infiltration  may  be  considerably  influenced  by  tem- 
perature difference  or  chimney  effect  which  will  operate  to  produce  a 
head  that  will  add  to  the  effect  of  the  wind  at  lower  levels  and  subtract 
from  it  at  higher  levels.2  On  the  other  hand,  the  wind  velocity  at  lower 
levels  may  be  somewhat  abated  by  surrounding  obstructions.  Further- 
more, the  chimney  effect  is  reduced  in  multi-story  buildings  by  the  partial 
isolation  of  floors  preventing  free  upward  movement,  so  that  wind  and 
temperature  difference  may  seldom  cooperate  to  the  fullest  extent. 
Making  the  rough  assumption  that  the  neutral  zone  is  located  at  mid- 


*Influence  of  Stack  Effect  on  the  Heat  Loss  in  Tall  Buildings,  by  Axel  Marin  (A.S.H.V.E.  Journal 
Section,  Heating,  Piping  and  Air  Conditioning*  August,  1934,  p.  349). 

126 


CHAPTER  6 — AIR  LEAKAGE 


height  of  a  building,  and  that  the  temperature  difference  is  70  F,  the 
following  formulae  may  be  used  to  determine  an  equivalent  wind  velocity 
to  be  used  in  connection  with  Tables  1  and  2  that  will  allow  for  both  wind 
velocity  and  temperature  difference: 


-  1.75  a  (1) 


-f-  1.75  b  (2) 

where 

Me  =  equivalent  wind  velocity  to  be  used  in  conjunction  with  Tables  1  and  2. 
M  =  wind  velocity  upon  which  infiltration  would  be  determined  if  tem- 
perature difference  were  disregarded. 
a  =  distance  of  windows  under  consideration  from  mid-height  of  building 

if  above  mid-height. 
b  =  distance  if  below  mid-height. 

The  coefficient  1.75  allows  for  about  one-half  the  temperature  difference  head. 

For  buildings  of  unusual  height,  Equation  1  would  indicate  negative 
infiltration  at  the  highest  stories,  which  condition  may,  at  times,  actually 
exist,  although  probably  no  greater  wind  velocities  should  be  figured  at 
such  extremely  high  levels3. 

Sealing  of  Vertical  Openings4 

In  tall,  multi-story  buildings,  every  effort  should  be  made  to  seal  off 
vertical  openings  such  as  stair-wells  and  elevator  shafts  from  the  re- 
mainder of  the  building.  Stair-wells  should  be  equipped  with  self-closing 
doors,  and  in  exceptionally  high  buildings,  should  be  closed  off  into 
sections  of  not  over  10  floors  each.  Plaster  cracks  should  be  filled. 
Elevator  enclosures  should  be  tight  and  solid  doors  should  be  used. 

If  'the  sealing  of  the  vertical  openings  is  made  effective,  no  allowance 
need  be  made  for  the  chimney  effect.  Instead,  the  greater  wind  move- 
ment at  the  high  altitudes  makes  it  advisable  to  install  additional  heating 
surface  on  the  upper  floors  above  the  level  of  neighboring  buildings,  this 
additional  surface  being  increased  as  the  height  is  increased.  One 
arbitrary  rule  is  to  increase  the  heating  surface  on  floors  above  neighboring 
buildings  by  an  amount  ranging  from  5  per  cent  to  20  per  cent.  This  extra 
heating  surface  is  required  only  on  the  windward  side  and  on  windy  days, 
and  hence  automatic  temperature  control  is  especially  desirable  with  such 
installations. 

Heating  Surface  for  Stair- Wells4 

In  stair-wells  that  are  open  through  many  floor  levels  although  closed 
off  from  the  remainder  of  each  floor  by  doors  and  partitions,  the  strati- 
fication of  air  makes  it  advisable  to  increase  the  amount  of  heating  surface 
at  the  lower  levels  and  to  decrease  the  amount  at  higher  levels  even  to  the 
point  of  omitting  all  heating  surface  on  the  top  several  floor  levels.  One 
rule  is  to  calculate  the  heating  surface  of  the  entire  stair- well  in  the  usual 


3Wind  Velocities  Near  a  Building  and  Their  Effect  on  Heat  Loss,  by  F.  C.  Houghten,  J.  L.  Blackshaw, 
and  Carl  Gutberlet  (A.S.H,V.E.  Journal  Section,  Healing,  Piping  and  Air  Conditioning,  September,  1934). 
*See  Flue  Action  in  Tall  Buildings,  by  H.  L.  Alt  (Heating,  Piping  and  Air  Conditioning,  May,  1932). 

127 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

way  and  to  place  50  per  cent  of  this  in  the  bottom  third,  the  normal 
amount  in  the  middle  third  and  the  balance  in  the  top  third. 

HEAT  EQUIVALENT  OF  AIR  ENTERING  BY  INFILTRATION 

The  heat  required  to  warm  cold,  outside  air,  which  enters  a  room  by 
infiltration,  to  the  temperature  of  the  room  is  given  by  the  following 
equation : 

Hi  «  0.24  Q  d  (t  -  t0)  (3) 

where 

Hi  =  Btu  per  hour  required  for  heating   air   leaking   Into    building   from 
outside  temperature  t0  to  inside  temperature  /. 

Q  —  cubic  feet  of  air  entering  per  hour  at  inside  temperature  /. 
d  =  density  (pounds  per  cubic  foot)  of  air  at  inside  temperature  t. 
t  =  inside  temperature  at  the  proper  level. 

t0  —  outside  air  temperature  for  which  heating  system  is  designed. 
0.24  =  specific  heat  of  air. 

It  is  sufficiently  accurate  to  take  d  —  0.075  Ib,  in  which  case  the  equa- 
tion reduces  to 

Hi  =  0.018  Q(t-  t0)  (4) 

While  a  heating  reserve  must  be  provided  to  warm  inleaking  air  on 
the  windward  side  of  a  building,  this  does  not  necessarily  mean  that  the 
heating  plant  must  be  provided  with  a  reserve  capacity,  since  the  inleaking 
air,  warmed  at  once  by  adequate  heating  surface  in  exposed  rooms,  will 
move  transversely  and  upwardly  through  the  building,  thus  relieving 
other  radiators  of  a  part  of  their  load.  The  actual  loss  of  heat  of  a  building 
caused  by  infiltration  is  not  to  be  confused  with  the  necessity  for  pro- 
viding additional  heating  capacity  for  a  given  space.  Infiltration  is  a 
disturbing  factor  in  the  heating  of  a  building,  and  its  maximum  effect 
(maximum  in  the  sense  of  an  average  of  wind  velocity  peaks  during  the 
heating  season  above  some  reasonably  chosen  minimum)  must  be  met 
by  a  properly  distributed  reserve  of  heating  capacity,  which  reserve,  how- 
ever, is  not  in  use  at  all  places  at  the  same  time,  nor  in  any  one  place  at 
all  times. 

REFERENCES 

Air  Leakage,  by  Houghten  and  Schrader  (A.S.H.V.E.  TRANSACTIONS,  Vol.  30,  1924). 

Air  Infiltration  through  Various  Types  of  Brick  Wall  Construction,  by  Larson,  Nelson  and  Braatz 
(A.S.H.V.E.  TRANSACTIONS,  Vol. -85,  1929). 

Infiltration  through  Plastered  and  Unplastered  Brick  Walls,  by  F.  C.  Houghten  and  Margaret  Ingels 
(A.S.H.V.E.  TRANSACTIONS,  Vol.  33,  1927). 

Air  Leakage  around  Window  Openings,  by  C.  C-  Schrader  (A.S.H.V.E.  TRANSACTIONS,  Vol.  30,  1924). 

Effect  of  Frame  Calking  and  Storm  Sash  on  Infiltration  around  and  through  Windows,  by  Richtrnann 
and  Braatz  (A.S.H.V.E.  TRANSACTIONS,  Vol.  34,  1928). 

Air  Leakage  on  Metal  Windows  in  a  Modern  Office  Building,  by  Houghten  and  O'Connell  (A.S.H.V.E, 
TRANSACTIONS,  Vol.  34,  1928). 

'  The  Weathertightness  of  Rolled  Section  Steel  Windows,  by  Emswiler  and  Randall  (A.S.H.V.E.  TRANS- 
ACTIONS, Vol.  34,  1928). 

Air  Leakage  through  a  Pivoted  Metal  Window,  by  Houghten  and  O'Connell  (A.S.H.V.E.  TRANSACTIONS, 
Vol.  34,  1928). 

Pressure  Difference  across  Windows  in  Relation  to  Wind  Velocity,  by  Emswiler  and  Randall  (A.S.H.V.E. 
TRANSACTIONS,  Vol.  35,  1929). 

128 


CHAPTER  6 — AIR  LEAKAGE 


Air  Infiltration  Through  Various  Types  of  Wood  Frame  Construction,  by  Larson,  Xelson  and  Braatz 
(A.S.H.V.E.  TRANSACTIONS,  Vol.  36,  1930). 

Neutral  Zone  in  Ventilating,  by  J.  E.  Emswiler  (A.S.H.V.E.  TRANSACTIONS,  Vol.  32,  1920). 

Air  Infiltration  Through  Double-Hung  Wood  Windows,  by  Larson,  Nelson  and  Kubasta  (A.S.H.V.E. 
TRANSACTIONS,  Vol.  37,  1931). 

Flue  Action  in  Tall  Buildings,  by  H.  L.  Alt  (Heating,  Piping  and  Air  Conditioning,  May,  1932;. 

Air  Infiltration  Through  Steel  Framed  Windows,  by  D.  O.  Rusk,  V.  H.  Cherry  and  L.  Boelter  (Heating, 
Piping  and  Air  Conditioning,  October,  1932). 

Investigation  of  Air  Outlets  in  Class  Room  Ventilation,  by  Larson,  Nelson,  and  Kubasta  (A.S.H.V.E. 
TRANSACTIONS,  Vol.  38,  1932). 

PROBLEMS  IN  PRACTICE 

1  •  What  are  the  causes  of  infiltration  (or  exfiltration)  and  how  do  they  act 
on  a  building? 

The  wind  and  temperature  differences  create  differences  between  internal  and  external 
pressures  which  cause  air  to  flow  through  any  openings  in  the  walls. 

2  •  Why  is  it  essential  to  consider  this  in  heating  calculations? 

The  inflowing  air  displaces  inside  heated  air  and  must  be  heated  up  to  the  internal 
temperature. 

3  •  Where   is   it   necessary    to   consider   infiltration   created   by   temperature 
difference? 

In  tall,  single-story  buildings  and  in  multi-story  buildings  where  the  floors  are  not 
adequately  isolated. 

4  •  Why  is  the  infiltration  in  a  building  less  than  that  determined  in  laboratory 
tests? 

In  laboratory  tests,  the  indicated  wind  velocity  is  measured  by  the  difference  in  pressure 
on  the  two  sides  of  a  single  wall,  window,  or  object  tested.  In  a  building,  an  internal 
back  pressure  is  built  up  between  its  walls  to  a  point  where  outflow  on  the  lee  side  is  equal 
to  inflow  on  the  windward  side  and  this  back  pressure  reduces  the  actual  inflow  below 
that  determined  in  the  laboratory  for  a  comparable  wind. 

5  •  Is  heat  loss  by  infiltration  through  walls  of  importance? 

Only  in  the  case  of  simple  walls  or  poorly  constructed  compound  walls. 

6  •  What  measurements  are  required  to  calculate  the  heat  loss  through  double- 
hung  wood  windows? 

Sash  crack  (equal  to  the  sash  perimeter  plus  the  meeting  rail)  and  frame  crack  (equal 
to  the  frame  perimeter). 

7  •  What  is  the  basis  for  selecting  the  wind  velocity  and  outside  temperature 
to  be  used  in  making  infiltration  calculations? 

Weather  Bureau  records.  The  wind  velocity  taken  is  the  average  during  the  three 
coldest  months  and  the  temperature  used  is  the  lowest  on  record  for  the  given  locality. 

8  •  How  does  the  temperature  difference  influence   the  heat   loss   in   a    tali 
building? 

The  chimney  effect  caused  by  the  temperature  difference  operates  to  produce  a  head  that 
will  add  to  the  effect  of  the  wind  at  lower  levels  and  subtract  from  it  at  higher  levels. 

9  •  For  a  wind  velocity  of  15  mph  and  a  building  180  ft  high,  calculate  the 
effective  wind  velocity  at  the  ground  floor  and  at  a  height  of  150  ft. 

a.  At  the  ground  floor  the  effective  wind  velocity  would  be 

Me  =  Vl52  +  1.75  X  90  =  19.6  mph 
129 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 
b.  At  a  floor  150  ft  above  the  ground 


Me  =  Vl52  -  175  X  60  =  11.0  mph 

10  •  A  room  contains  three  2  ft -8  in.  by  5  ft-6  in.  plain  double-hung  wood  win- 
dows with  Jle-hi.   crack  and  %4-in.  clearance.     Assume  a  wind  velocity  of 
20  mph  and  a  temperature  difference  of  75  F.    Neglecting  chimney  effect,  what 
is  the  maximum  heat  loss  due  to  infiltration? 

From  Table  2,  heat  loss  per  foot  of  crack  per  degree  temperature  difference  is  1.067  Btu 
per  hour.  Length  of  crack  for  the  three  windows  is  57  ft.  The  maximum  heat  loss,  due 
to  infiltration,  is  equal  to  1.067  X  57  X  75  or  4561  Btu  per  hour. 

11  •  Find  the  infiltration  through  a  wall  with  16-in.  shingles  on  1  in.  by  4  in. 
hoards  w^th  20  mph  wind  velocity.    Give  the  pressure  drop  through  the  wall. 

Referring  to  Curve  3C,  Fig.  1,  the  value  on  the  horizontal  scale  corresponding  to  20  mph 

is  approximately  102  cfh  per  square  foot  of  wall. 

The  pressure  drop  through  the  wall  is  0.193  in.  of  water  (see  left  hand  vertical  scale). 

12  •  What  will  be  the  infiltration  through  air-dried   end   and    side-matched 
sheathing  for  15  mph  wind  velocity? 

Referring  to  Curve  IOC,  Fig.  2,  the  value  on  the  horizontal  scale  corresponding  to 
15  mph  is  50  cfh  per  square  foot  of  wall. 

13  •  From  Table  2,  find  the  infiltration  (cubic  feet  per  hour  per  foot  of  crack) 
for  an  average  double-hung  window,  not  weather  stripped,  with  a  20  mph 
wind  velocity. 

59.3  cu  ft  per  foot  of  crack  per  hour. 

14  •  Using  the  value  found  in  Question  11,  what  will  be  the  heat  requirement 
in  a  building  with  a  total  crack  (all  windows  and  doors)  of  180  ft  if  the  wind 
velocity  is  15  mph,  the  outside  temperature  is  0  F,  and  the  inside  temperature 
is  70  F? 

Using  one  half  of  the  total  crack,  the  volume  of  air  is: 
90  X  59.3  =  5337  cu  ft 
H  -  0.018  X  5337  X  (70  -  0)  =  6724.6  Btu.    (See  Equation  4.) 


130 


Chapter  7 

HEATING  LOAD 


Factors     Governing    Heat    Demand,     Procedure,     Temperatures, 

Wind    Movement,    Heat    Sources    Other    Than    Heating    Plant, 

Example,  Condensation 


design  any  system  of  heating,  the  maximum  probable  heat  demand 
JL  must  be  accurately  estimated  in  order  that  the  apparatus  installed 
shall  be  capable  of  maintaining  the  desired  temperature  at  all  times.  The 
factors  which  govern  this  maximum  heat  demand  —  most  of  which  are 
seldom,  if  ever,  in  equilibrium  —  include  the  following: 


1.  Outside  temperature. 

2.  Rain  or  snow. 

3.  Sunshine  or  cloudiness. 

4.  Wind  velocity. 

5.  Heat  transmission  of  exposed  parts  of  building. 

6.  Infiltration  of  air  through  cracks,  crevices  and 

open  doors  and  windows. 

7.  Heat  capacity  of  materials. 

8.  Rate  of  absorption  of  solar  radiation  by  exposed 

materials. 

9.  Inside  temperatures. 

10.  Stratification  of  air. 

11.  Type  of  heating  system. 

12.  Ventilation  requirements. 

13.  Period  and  nature  of  occupancy. 

14.  Temperature  regulation. 


Outside  Conditions 
(The  Weather} 


Building 
Construction 


Inside 
Conditions 


The  inside  conditions  vary  from  time  to  time,  the  physical  properties  of 
the  building  construction  may  change  with  age,  and  the  outside  conditions 
are  changing  constantly.  Just  what  the  worst  combination  of  all  of  these 
variable  factors  is  likely  to  be  in  any  particular  case  is  therefore  con- 
jectural. Because  of  the  nature  of  the  problem,  extreme  precision  in 
estimating  heat  losses  at  any  time,  while  desirable,  is  hard  of  attainment. 

The  procedure  to  be  followed  in  determining  the  heat  loss  from  any 
building  can  be  divided  into  seven  consecutive  steps,  as  follows: 

1.  Determine  on  the  inside  air  temperature,  at  the  breathing  line  or  the  30-in.  line, 
which  is  to  be  maintained  in  the  building  during  the  coldest  weather.  (See  Table  1.) 

2.  Determine  on  an  outside  air  temperature  for  design  purposes,  based  on  the  minimum 
temperatures  recorded  in  the  locality  in  question,  which  will  provide  for  all  but  the 
most  severe  weather  conditions.   Such  conditions  as  may  exist  for  only  a  few  consecu- 
tive hours  are  readily  taken  care  of  by  the  heat  capacity  of  the   buHtKng  Itself. 
(See  Table  2.) 

131 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

3  Select  or  compute  the  heat  transmission  coefficients  for  outside  walls  and  glass; 
also  for  inside  walls,  floors,  or  top-floor  ceilings,  if  these  are  next  to  unheated  space; 
include  roof  if  next  to  heated  space.    (See  Chapter  5.) 

4  Measure  up  net  outside  wall,  glass  and  roof  next  to  heated  spaces,  as  well  as  any 
cold  walls,  floors  or  ceilings  next  to  unheated  space.  Such  measurements  are  made  from 
building  plans,  or  from  the  actual  building. 

5.  Compute  the  heat  transmission  losses  for  each  kind  of  wall,  glass,  floor,  ceiling 
and  roof  in  the  building  by  multiplying  the  heat  transmission  coefficient  in  each  case 
by  the  area  of  the  surface  in  square  feet  and  the  temperature  difference  between  the 
inside  and  outside  air.    (See  Items  1  and  2.) 

6.  Select  unit  values  and  compute  the  heat  equivalent  of  the  infiltration  of  cold  air 
taking  place  around  outside  doors  and  windows.  These  unit  values  depend  on  the  kind  or 
width  of  crack  and  wind  velocity,  and  when  multiplied  by  the  length  of  crack  and  the 
temperature  difference  between  the  inside  and  outside  air,  the  result  expresses  the  heat 
required  to  warm  up  the  cold  air  leaking  into  the  building  per  hour,    (bee  Chapter  b.) 

7.  The  sum  of  the  heat  losses  by  transmission  (Item  5)  through  the  outside  wall  and 
glass  as  well  as  through  any  cold  floors,  ceilings  or  roof,  plus  the  heat  equivalent  (Item  6) 
of  the  cold  air  entering  by  infiltration  represents  the  total  heat  loss  equivalent  for  any 
building. 

Item  7  represents  the  heat  losses  after  the  building  is  heated  and  under 
stable  operating  conditions  in  coldest  weather.  Additional  heat  is 
required  for  raising  the  temperature  of  the  air,  the  building  materials  and 
the  material  contents  of  the  building  to  the  specified  standard  inside 
temperature. 

The  rate  at  which  this  additional  heat  is  required  depends  upon  the 
heat  capacity  of  the  structure  and  its  material  contents  and  upon  the 
time  in  which  these  are  to  be  heated. 

This  additional  heat  may  be  figured  and  allowed  for  as  conditions  re- 


TABLE  1.    WINTER  INSIDE  DRY-BULB  TEMPERATURES  USUALLY  SPECiFiEDa 


TYPE  OP  BUILDING 


DEO  FAER 


TYPE  OP  BUILDING 


DBG  F^ 


SCHOOLS 

Class  rooms 

Assembly  rooms 

Gymnasiums 

Toilets  and  baths 

Wardrobe  and  locker  rooms  „ 

Kitchens 

Dining  and  lunch   rooms 

Playrooms 

Natatoriums™ 


HOSPITALS — 

Private  rooms 

Private  rooms  (surgical) 

Operating  rooms 

Wards - 

Kitchens  and  laundries 

Toilets 

Bathrooms 


70-72 
68-72 
55-65 

70 
65-68 

66 

65-70 
60-65 

75 


70-72 
70-80 
70-95 

68 

66 

68 
70-80 


THEATERS — 
Seating  space.. 
Lounge  rooms.. 
Toilets 


HOTELS — 

Bedrooms  and  baths 

Dining  rooms 

Kitchens  and  laundries.... 

Ballrooms 

Toilets  and  service  rooms.. 


HOMES - 

STORES 

PUBLIC  BUILDINGS- 

WARM  AIR  BATHS 

STEAM  BATHS 

FACTORIES  AND  MACHINE  SHOPS 

FOUNDRIES  AND  BOILER  SHOPS 

PAINT  SHOPS 


68-72 

68-72 

68 


70 

70 

66 

65-68 


70-72 

65-68 

68-72 

120 

110 

60-65 

50-60 

80 


aThe  most  comfortable  dry-bulb  temperature  to  be  maintained  depends  on  the  relative  humidity  and 
air  motion.  These  three  factors  considered  together  constitute  what  ts  termed  the  effective  temperature. 
See  Chapter  2. 

132 


CHAPTER  7 — HEATING  LOAD 


quire,  but  inasmuch  as  the  heating  system  proportioned  for  taking  care 
of  the  heat  losses  will  usually  have  a  capacity  about  100  per  cent  greater 
than  that  required  for  average  winter  weather,  and  inasmuch  as  most 
buildings  may  either  be  continuously  heated  or  have  more  time  allowed 
for  heating-up  during  the  few  minimum  temperature  days,  no  allowance 
is  made  except  in  the  size  of  boilers  or  furnaces. 

INSIDE  TEMPERATURES 

The  inside  air  temperature  which  must  be  maintained  within  a  building 
and  which  should  always  be  stated  in  the  heating  specifications  is  under- 
stood to  be  the  dry-bulb  temperature  at  the  breathing  line,  5  ft  above  the 
floor,  or  the  30-in.  line,  and  not  less  than  3  ft  from  the  outside  walls. 
Inside  air  temperatures,  usually  specified,  vary  in  accordance  with  the  use 
to  which  the  building  is  to  be  put  and  Table  1  presents  values  which  con- 
form with  good  practice. 

The  proper  dry-bulb  temperature  to  be  maintained  depends  upon  the 
relative  humidity  and  air  motion,  as  explained  in  Chapter  2.  In  other 
words,  a  person  may  feel  warm  or  cool  at  the  same  dry-bulb  temperature, 
depending  on  the  relative  humidity  and  air  motion.  The  optimum  winter 
effective  temperature  for  sedentary  persons,  as  determined  at  the  A.S.H. 
V.E.  Research  Laboratory,  is  66  deg.1 

According  ^ to  Fig.  2,  Chapter  2,  for  so-called  still  air  conditions,  a 
relative  humidity  of  approximately  50  per  cent  is  required  to  produce  an 
effective  temperature  of  66  deg  when  the  dry-bulb  temperature  is  70  F. 
However,  even  where  provision  is  made  for  artificial  humidification,  the 
relative  humidity  is  seldom  maintained  higher  than  40  per  cent  during  the 
extremely  cold  weather,  and  where  no  provision  is  made  for  humidifica- 
tion, the  relative  humidity  may  be  20  per  cent  or  less.  Consequently,  in 
using  the  figures  given  in  Table  1,  consideration  should  be  given  to 
whether  provision  is  to  be  made  for  humidification,  and  if  so,  the  actual 
relative  humidity  to  be  maintained. 

Temperature  at  Proper  Level:  In  making  the  actual  heat-loss  compu- 
tations, however,  for  the  various  rooms  in  a  building  it  is  often  necessary 
to  modify  the  temperatures  given  in  Table  1  so  that  the  air  temperature 
at  the  proper  level  will  be  used.  By  air  temperature  at  the  proper  level  is 
meant,  in  the  case  of  walls,  the  air  temperature  at  the  mean  height  be- 
tween floor  and  ceiling;  in  the  case  of  glass,  the  air  temperature  at  the 
mean  height  of  the  glass;  in  the  case  of  roof  or  ceiling,  the  air  temperature 
at  the  mean  height  of  the  roof  or  ceiling  above  the  floor  of  the  heated 
room; and  in  the  case  of  floors,  the  air  temperature  at  the  floor  level.  In 
the  case  of  heated  spaces  adjacent  to  unheated  spaces,  it  will  usually  be 
sufficient  to  assume  the  temperature  in  such  spaces  as  the  mean  between 
the  temperature  of  the  inside  heated  spaces  and  the  outside  air  tempera- 
ture, excepting  where  the  combined  heat  transmission  coefficient  of  the 
roof  and  ceiling  can  be  used,  in  which  case  the  usual  inside  and  outside 
temperatures  should  be  applied.  (See  discussion  regarding  the  use  of 
combined  coefficients  of  pitched  roofs,  unheated  attics  and  top-floor 
ceilings  Chapter  5.) 


*See  Chapter  2,  p.  43. 

133 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

High  Ceilings:  Research  data  concerning  stratification  of  air  in  build- 
ings are  lacking,  but  in  general  it  may  be  said  that  where  the  increase  in 
temperature  is  due  to  the  natural  tendency  of  the  warmer  or  less  dense 
air  to  rise,  as  where  a  direct  radiation  system  is  installed,  the  temperature 
of  the  air  at  the  ceiling  increases  with  the  ceiling  height.  The  relation, 
however,  is  not  a  straight-line  function,  as  the  amount  of  increase  per  foot 
of  height  apparently  decreases  as  the  height  of  the  ceiling  increases,  ac- 
cording to  present  available  information. 

Where  ceiling  heights  are  under  20  ft,  it  is  common  engineering  practice 
to  consider  that  the  Fahrenheit  temperature  increases  2  per  cent  for  each 
foot  of  height  above  the  breathing  line.  This  rule,  sufficiently  accurate 
for  most  cases,  will  give  the  probable  air  temperature  at  any  given  level 
for  a  room  heated  by  direct  radiation.  Thus,  the  probable  temperature 
in  a  room  at  a  point  three  feet  above  the  breathing  line,  if  the  breathing 
line  temperature  is  70  F,  will  be 

(1.00  +  3  X  .02)  70  =  74.2  F. 

With  certain  types  of  heating  and  ventilating  systems,  which  tend  to 
oppose  the  natural  tendency  of  warm  air  to  rise,  the  temperature  differ- 
ential between  floor  and  ceiling  can  be  greatly  reduced.  These  include 
unit  heaters,  fan-furnace  heaters,  and  the  various  types  of  mechanical 
ventilating  systems.  The  amount  of  reduction  is  problematical  in  certain 
instances,  as  it  depends  upon  many  factors  such  as  location  of  heaters, 
air  temperature,  and  direction  and  velocity  of  air  discharge.  In  some 
cases  it  has  been  possible  to  reduce  the  temperature  between  the  floor 
and  ceiling  by  a  few  degrees,  whereas,  in  other  cases,  the  temperature  at 
the  ceiling  has  actually  been  increased  because  of  improper  design,  instal- 
lation or  operation  of  equipment.  So  much  depends  upon  the  factors 
enumerated  that  it  is  not  advisable  to  allow  less  than  1  per  cent  per  foot 
(and  usually  more)  above  the  breathing  line  in  arriving  at  the  air  tem- 
perature at  any  given  level  for  any  of  these  types  of  heating  and  ventilating 
systems,  unless  the  manufacturers  are  willing  to  guarantee  that  the  par- 
ticular type  of  equipment  under  consideration  will  maintain  a  smaller 
temperature  differential  for  the  specific  conditions  involved. 

Temperature  at  Floor  Level:  In  determining  mean  air  temperatures 
just  above  floors  which  are  next  to  ground  or  unheated  spaces,  a  tempera- 
ture 5  deg  lower  than  the  breathing-line  temperature  may  be  used,  pro- 
vided the  breathing-line  temperature  is  not  less  than  55  F. 

OUTSIDE  TEMPERATURES 

The  outside  temperature  used  in  computing  the  heat  loss  from  a  build- 
ing is  seldom  taken  as  the  lowest  temperature  ever  recorded  in  a  given 
locality.  Such  temperatures  are  usually  of  short  duration  and  are  rarely 
repeated  in  successive  years.  It  is  therefore  evident  that  a  temperature 
somewhat  higher  than  the  lowest  on  record  may  be  properly  assumed  in 
making  the  heat-loss  computations. 

The  outside  temperature  to  be  assumed  in  the  design  of  any  heating 
system  is  ordinarily  not  more  than  15  deg  above  .the  lowest  recorded  tem- 
perature as  reported  by  the  Weather  Bureau  during  the  preceding  10 
years  for  the  locality  in  which  the  heating  system  is  to  be  installed.  In 

134 


CHAPTER  7 — HEATING  LOAD 


the  case  of  massive  and  well  insulated  buildings  in  localities  where  the 
minimum  does  not  prevail  for  more  than  a  few  hours,  or  where  the  lowest 
recorded  temperature  is  extremely  unusual,  more  than  15  deg  above  the 
minimum  may  be  allowed,  due  primarily  to  the  fly -wheel  effect  of  the  heat 
capacity  of  the  structure.  The  outside  temperature  assumed  and  used  in 
the  design  should  always  be  stated  in  the  heating  specifications.  Table  2 
lists  the  coldest  dry-bulb  temperatures  ever  recorded  by  the  Weather 
Bureau  at  the  places  listed. 

If  Weather  Bureau  reports  are  not  available  for  the  locality  in  question, 
then  the  reports  for  the  station  nearest  to  this  locality  are  to  be  used, 
unless  some  other  temperature  is  specifically  stated  in  the  specifications. 
In  computing  the  average  heat  transmission  losses  for  the  heating  season 
in  the  United  States  the  average  outside  temperature  from  October  1 
to  May  1  should  be  used. 

WIND  MOVEMENT 

Trie  effect  of  wind  on  the  heating  requirements  of  any  building  should 
be  given  consideration  under  two  heads: 

1.  Wind  movement  increases  the  heat  transmission  of  walls,  glass,  and  roof,  affecting 
poor  walls  to  a  much  greater  extent  than  good  walls. 

2.  Wind  movement  materially  increases  the  infiltration  (inleakage)  of  cold  air  through 
the  cracks  around  doors  and  windows,  and  even  through  the  building  materials  them- 
selves, if  such  materials  are  at  all  porous. 

Theoretically  as  a  basis  for  design,  the  most  unfavorable  combination 
of  temperature  and  wind  velocity  should  be  chosen.  It  is  entirely  possible 
that  a  building  might  require  more  heat  on  a  windy  day  with  a  moderately 
low  outside  temperature  than  on  a  quiet  day  with  a  much  lower  outside 
temperature.  However,  the  combination  of  wind  and  temperature  which 
is  the  worst  would  differ  with  different  buildings,  because  wind  velocity 
has  a  greater  effect  on  buildings  which  have  relatively  high  infiltration 
losses.  It  would  be  possible  to  work  out  the  heating  load  for  a  building 
for  several  different  combinations  of  temperature  and  wind  velocity  which 
records  show  to  have  occurred  and  to  select  the  worst  combination ;  but 
designers  generally  do  not  feel  that  such  a  degree  of  refinement  is  justified. 
Therefore,  pending  further  studies  of  actual  buildings,  it  is  recommended 
that  the  average  wind  movement  in  any  locality  during  December, 
January  and  February  be  provided  for  in  computing  (1)  the  heat  trans- 
mission of  a  building,  and  (2)  the  heat  required  to  take  care  of  the  infiltra- 
tion of  outside  air. 

The  first  condition  is  readily  taken  care  of,  as  explained  in  Chapter  5, 
by  using  a  surface  coefficient /0  for  the  outside  wall  surface  which  is  based 
on  the  proper  wind  velocity.  In  case  specific  data  are  lacking  for  any 
given  locality,  it  is  sufficiently  accurate  to  use  an  average  wind  velocity  of 
approximately  15  mph  which  is  the  velocity  upon  which  the  heat  trans- 
mission coefficient  tables  in  Chapter  5  are  based. 

In  a  similar  manner,  the  heat  allowance  for  infiltration  through  cracks 
and  walls  (Tables  1  and  2,  Chapter  6)  must  be  based  on  the  proper  wind 
velocity  for  a  given  locality.  In  the  case  of  tall  buildings  special  attention 
must  be  given  to  infiltration  factors.  (See  Chapter  6). 

In  the  past  many  designers  have  used  empirical  exposure  factors  which 

135 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


TABLE  2.    CLIMATIC  CONDITIONS  COMPILED  FROM  WEATHER  BUREAU  RECORDS 


COL.  A 

COL.  B 

COL.  C 

COL.  D 

COL.  E 

COL.  F 

State 

City 

Average 
Temp., 
Oct.  1st- 
May  1st 

Lowest 
Tempera- 
ture 
Ever 
Reported 

Average 
Wind  Vel- 
ocity Dec., 
Jan.,  Feb.. 
Miles  per 
Hour 

Direction 
of  Prevail- 
ing Wind, 
Dec..  Jan., 
Feb. 

Ala 

Mobile  

57.7 

^ 

8.3 

N 

Birmingham 

53.9 

-10 

8.6 

N 

Ariz 

Phoenix. 

59.5 

12 

3.9 

E 

Flagstaff 

34.9 

—25 

6.7 

SW 

Ark 

Fort  Smith 

49.5 

—15 

8.0 

E 

Calif 

Little  Rock.  
San  Francisco  

51.6 
54.3 

-12 

27 

9.9 

7.5 

NW 
N 

Los  Ansreles            .               

58.6 

28 

6.1 

NE 

Colo 

"                    o**"**    .—  ...                 ......  ...     . 

Denver 

39.3 

—29 

7.4 

s 

Conn 

Grand  Junction  
New  Haven 

39.2 
38.0 

-21 
-15 

5.6 
9.3 

SE 

N 

D.  C. 

Washington  ...          .  _ 

43.2 

-15 

7.3 

NW 

Fla 

Jacksonville 

61.9 

10 

8.2 

NE 

Ga 

Atlanta    ...... 

51.4 

—8 

11.8 

NW 

Idaho  .. 

Savannah  
Lewiston  „ 

58.4 
42.5 

8 
-23 

8.3 

4.7 

NW 
E 

Pocatello 

36.4 

—22 

9.3 

SE 

Ill  

Chicago  
Springfield  

36.4 
39.9 

-23 
-24 

17.0 
10.2 

SW 

NW 

Ind  

Indianapolis  

Evansville 

40.2 
44.1 

-25 

—  16 

11.8 

8.4 

S 

s 

Iowa  

Kans. 

Dubuque  
Sioux  City  

Concordia  

33.9 
32.1 
38.9 

-32 
-35 
-25 

6.1 
12.2 
7.3 

NW 
NW 
N 

Dodge  Citv 

40.2 

—26 

10.4 

NW 

Ky.        ~    . 
La. 

^  v  .»^  .„    y  —  
Louisville  

New  Orleans 

45.2 
61.5 

-20 

7 

9.3 
9.6 

SW 

N 

Me  

Md. 

Shreveport  
Eastport  .  
Portland  

Baltimore 

56.2 
31.1 
33.6 
43.6 

-5 
-23 

-21 

7 

7.7 
13.8 
10.1 
7.2 

SE 
W 

NW 
NW 

Mass.  

Mich. 

Boston  . 

Alpena.    . 

37.6 
29.1 

-18 

-28 

11.7 
11.3 

W 
W 

Detroit-  

Marquette 

35.4 
27.6 

-24 
—27 

13.1 
11.4 

SW 

NW 

Minn.  

Duluth  

Minneapolis 

25.1 
29.6 

-41 
—33 

11.1 
11.5 

SW 

NW 

Miss  

Mo. 

Vicksburg  

St.  Joseph 

56.0 
40.3 

-1 
—24 

7.6 
9.1 

SE 

NW 

St.  Louis  

43.3 

—22  ' 

11.8 

NW 

Springfield 

43.0 

—29 

11.3 

SE 

Mont  

Billings  

34.7 

-49 

W 

Havre 

27.7 

—57 

8.7 

SW 

Nebr  

Lincoln  

37.0 

-29 

10.9 

N 

North  Platte  

34.6 

-35 

9.0 

W 

Nev 

Tonopah 

39  6 

—  10 

9  9 

SE 

Winnemucca  

37.9 

-28 

9.5 

NE 

N.  H.._  

Concord  _ 

33.4 

-35 

6.0 

NW 

N.  J. 

Atlantic  City 

41.6 

—9 

10.6 

NW 

N.Y..Z1L.. 

Albany  
Buffalo  

35.1 
34.7 

-24 

-20 

7.9 
17.7 

S 
W 

N.  M  

New  York.  
Santa  Fe  _  

40.7 
38.0 

-14 
-13 

17.1 
7.3 

NW 
NE 

136 


CHAPTER  7 — HEATING  LOAD 


TABLE  2.    CLIMATIC  CONDITIONS  COMPILED  FROM  WEATHER  BUREAU  RECORDS — 

(Continued) 


COL.  A 

COL.  B 

COL.  C 

COL.  D 

COL.  E 

COL.  F 

State 
or 
Province 

City 

Average 
Temp., 
Oct.  1st- 
May  1st 

Lowest 
Tempera- 
ture 
Ever 
Reported 

Average 
Wind  Vel- 
ocity Dec., 
Jan.,  Feb., 
Miles  per 
Hour 

Direction 
of  Prevail-, 
ing  Wind, 
Dec.,  Jan., 
Feb. 

N  C 

Raleigh 

49.7 

53.1 
24.5 
18.9 
36.9 
39.9 
48.0 
34.1 
45.9 
41.9 
40.8 
37.6 
56.9 
53.7 
28.1 
32.3 
47.0 
50.9 
53.0 
54.7 
60.7 
38.1 
40.0 
29.3 
49.1 
45.2 
47.4 
45.3 
37.5 
38.8 
41.9 
28.6 
31.2 
33.0 
31.0 
28.9 
23.3 
43.8 
41.7 
17.2 
27.1 
35.5 
32.5 
26.9 
21.6 
32.0 
30.1 
27.4 
24.4 
14.7 
1.6 

-2 
5 
-45 
-44 
-17 
-20 
-17 
-24 
-2 
-6 
'  -20 
-17 
7 
-2 
-43 
-34 
-16 
-9 
-2 
-8 
4 
-24 
-20 
-28 
2 
-7 
-3 
3 
-30 
-28 
-27 
-36 
-43 
-25 
-45 
-40 
-57 
-2 
2 
-46 
35 

7.3 

8.9 

liTi 

14.5 
9.3 
12.0 
6.0 
6.5 
11.0 
13.7 
14.6 
11.0 
8.0 
11.5 
7.5 
6.5 
9.6 
10.5 
11.0 
8.2 
8.9 
4.9 
12.9 
9.0 
5.2 
7.4 
9.1 
5.2 
4.8 
6.6 
12.8 
5.6 
11.7 
5.3 
3.0 
4.5 
8.9 
4.2 
12.4 
8.7 
13.0 

SW 

SW 

NW 
W 

sw 
sw 

N 
SE 
S 

NW 
NW 
NW 
N 
NE 
NW 
W 
SW 
NW 
NW 
NW 
N 
W 
SE 
S 
N 
NW 
S 
SE 
SW 
W 
S 
SW 
NW 
W 
NW 
NE 
W 
N 
E 
SW 
NW 
NW 

W 

SW 

NW 
SW 

sw 
sw 

N.  Dak.,..  
Ohio  

Okla 

Wilmington 

Bismarck.-  

Devils  Lake 

Cleveland  ~  

Columbus 

Oklahoma  City 

Oree 

Baker 

Pa  

Portland. 

Philadelphia  

R.  I  

Pittsburgh 

Providence  

S  C. 

Charleston 

S.  Dak  
Te/m  

Texas 

Columbia  

Huron  
Rapid  City  

Knoxville 

Memphis  -  

El  Paso 

Utah 

Fort  Worth.  _       

San  Antonio  
Modena  

Vt  

Salt  Lake  City 

Burlington  

Va 

Norfolk 

Wash  

Lynchburg  

Richmond 

Seattle.-  

W.  Va  

Spokane 

Elkins.    

Wis  

Parkersburg 

Green  Bay 

Wyo  

La  Crosse 

Milwaukee  

Sheridan 

Alta  

Lander  

Edmonton 

B.  C. 

Victoria  

Man  

Vancouver 

Winnipeg      

N.  B 

Fredericton  -. 

N.  S  

Yarmouth  

-12 
26 

Ont  

London 

P.  E.  I  

Que.. 

Ottawa  

-33 
-51 
-28 
-27 
-27 
-34 
-70 
-68 

7.5 

13".~5 
8.7 
15.4 
15.0 
3.2 

Pt.  Arthur  
Toronto  

C  harlotteto  wn 

Montreal.  

Sask  _ 

Quebec,  ,. 

Prince  Albert  

Yukon  

Dawson 

137 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

were  arbitrarily  chosen  to  increase  the  calculated  heat  loss  on  the  side  or 
sides  of  the  building  exposed  to  the  prevailing  winds.  It  is  also  possible 
to  differentiate  among  the  various  exposures  more  accurately  by  calcu- 
lating the  infiltration  and  transmission  losses  separately  for  the  different 
sides  of  the  building,  using  different  assumed  wind  velocities.  Recent 
investigations  indicate,  however,  that  the  wind  direction  indicated  by 
Weather  Bureau  instruments  does  not  always  correspond  with  the 
direction  of  actual  impact  on  the  building  walls,  due  to  deflection  by 
surrounding  buildings. 

Pending  the  time  when  the  lack  of  actual  test  data  is  remedied,  it  is 
recommended  that  no  differentiation  be  made  among  the  various  sides  of 
a  building  in  calculating  the  heat  losses.  It  should  be  remembered  that 
the  values  for  U  in  the  tables  in  Chapter  5  are  based  on  a  wind  velocity 
of  15  mph. 

The  Heating,  Piping  and  Air  Conditioning  Contractors  National  Associ- 
ation has  devised  a  method2  for  calculating  the  square  feet  of  equivalent 
direct  radiation  required  in  a  building.  This  method  makes  use  of  ex- 
posure factors  which  vary  according  to  the  geographical  location  and  the 
angular  situation  of  the  construction  in  question  in  reference  to  pre- 
vailing winds  and  the  velocity  of  them. 

HEAT  FROM  SOURCES  OTHER  THAN  HEATING  PLANT 

The  heat  supplied  by  persons,  lights,  motors  and  machinery  should 
always  be  ascertained  in  the  case  of  theaters,  assembly  halls,  and  in- 
dustrial plants,  but  allowances  for  such  heat  sources  must  be  made  only 
after  careful  consideration  of  all  local  conditions.  In  many  cases,  ^  these 
heat  sources  should  not  be  allowed  to  affect  the  size  of  the  installation  at 
all,  although  they  may  have  a  marked  effect  on  the  operation  and  con- 
trol of  the  system.  In  general,  it  is  safe  to  say  that  where  audiences  are 
involved,  the  heating  installation  must  have  sufficient  capacity  to  bring 
the  building  up  to  the  stipulated  inside  temperature  before  the  audience 
arrives.  In  industrial  plants,  quite  a  different  condition  exists,  and  heat 
sources,  if  they  are  always  available  during  the  period  of  human  occu- 
pancy, may  be  substituted  for  a  portion  of  the  heating  installation.  In 
no  case  should  the  actual  heating  installation  (exclusive  of  heat  sources) 
be  reduced  below  that  required  to  maintain  at  least  40  F  in  the  building. 

Motors  and  Machinery 

Motors  and  the  machinery  which  they  drive,  if  both  are  located  in  the 
room,  convert  all  of  the  electrical  energy  supplied  into  Jheat,  which  is 
retained  in  the  room  if  the  product  being  manufactured  is  not  removed 
until  its  temperature  is  the  same  as  the  room  temperature. 

If  power  is  transmitted  to  the  machinery  from  the  outside,  then  only 

the  heat  equivalent  of  the  brake  horsepower  supplied  is  used,    In  the 

^  ,.    ^          i  Motor  horsepower  vxOC/,~        A 

first  case  the  Btu  supplied  per  hour  =  Efficiency  of  motor  X  2,546,  and 

in  the  second  case  Btu  per  hour  =  bhp  X  2,546,  in  which  2,546  is  the 
Btu  equivalent  of  1  hp-hour.  In  high-powered  mills  this  is  the  chief 

2See  Standards  of  Heating,  Piping  and  Air  Conditioning  Contractors  National  Association. 

138 


CHAPTER  7 — HEATING  LOAD 


source  of  heating  and  it  is  frequently  sufficient  to  overheat  the  building 
even  in  zero  weather,  thus  requiring  cooling  by  ventilation  the  year 
round. 

The  heat  (in  Btu  per  hour)  from  electric  lamps  is  obtained  by  multi- 
plying the  watts  per  lamp  by  the  number  of  lamps  and  by  3.415.  One 
cubic  foot  of  producer  gas  gives  off  about  150  Btu  per  hour;  one  cubic 
foot  of  illuminating  gas  gives  off  about  535  Btu  per  hour;  and  one  cubic 
foot  of  natural  gas  gives  off  about  1000  Btu  per  hour.  A  Welsbach 
burner  averages  3  cu  ft  of  gas  per  hour  and  a  fish-tail  burner,  5  cu  ft 
per  hour.  For  information  concerning  the  heat  supplied  by  persons, 
see  Chapter  2. 

In  intermittently  heated  buildings,  besides  the  capacity  necessary 
to  care  for  the  normal  heat  loss  which  may  be  calculated  according  to 
customary  rules,  additional  capacity  should  be  provided  to  supply  the 
heat  necessary  to  warm  up  the  cold  material  of  the  interior  walls,  floors, 
and  furnishings.  Tests  have  shown  that  when  a  cold  building  has  had  its 
temperature  raised  to  about  60  F  from  an  initial  condition  of  about  0  F, 
the  heat  absorbed  from  the  air  by  the  material  in  the  structure  may  vary 
from  50  per  cent  to  150  per  cent  of  the  normal  heat  loss  of  the  building. 
It  is  therefore  necessary,  in  order  to  heat  up  a  cold  building  within  a 
reasonable  length  of  time,  to  provide  such  additional  capacity.  If  the 
interior  material  is  cold  when  people  enter  a  building,  the  radiation  of 
heat  from  the  occupants  to  the  cold  material  will  be  greater  than  is 
normal  and  discomfort  will  result.  (See  Chapter  2.) 

CONDENSATION  ON  BUILDING  SURFACES3 

Condensation  on  the  interior  surfaces  of  buildings  is  often  a  serious 
problem.  Water  dripping  from  a  ceiling  may  cause  irreparable  damage 
to  manufactured  articles  and  machinery.  It  often  results  in  short-cir- 
cuiting of  electric  power  and  lighting  systems,  necessitating  shut-downs 
and  incurring  costly  repairs.  It  also  causes  rotting  of  wood  roof  struc- 
tures, corrosion  of  metal  roofs,  and  spalling  and  disintegration  of  gypsum 
and  other  types  of  roof  decks  not  properly  protected. 

Condensation  is  caused  by  the  contact  of  the  warm  humid  air  in  a 
building  with  surfaces  below  the  dew-point  temperature,  and  can  be 
remedied  in  two  ways,  (1)  by  increasing  the  temperature  of  such  surfaces 
above  the  dew-point  temperature,  or  (2)  by  lowering  the  humidity. 

Dehumidification,  of  course,  is  not  advisable  where  a  high  relative 
humidity  is  necessary  for  manufacturing  processes.  Hence,  the^  only 
alternative  is  to  increase  the  surface  temperature  by  decreasing  the  inside 
surface  resistance.  This  can  be  accomplished  by  increasing  the  velocity 
of  air  passing  over  the  surface,  or  by  increasing  the  over-all  Resistance  of 
the  wall  or  roof  by  installing  a  sufficient  thickness  of  insulation. 

The  latter  method  is  generally  used,  and  the  thickness  of  insulation 
is  determined  by  ascertaining  the  amount  of  resistance  to  be  added  ^  to 
increase  the  temperature  of  the  interior  surface  above  ^  the  dew-point 
temperature  for  the  maximum  conditions  involved.  This  in  turn  is  based 
on  the  fundamental  principle  that  the  drop  in  temperature  is  proportional 
to  the  resistance.  See  Question  1  at  the  end  of  this  chapter. 

2See  Preventing  Condensation  on  Interior  Building  Surfaces,  by  Paul  D.  Close  (A.S.H.V.E.  TRANS- 
ACTIONS, Vol.  36,  1930). 

139 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 
EXAMPLE  OF  HEAT  LOSS  COMPUTATIONS 


Window 
Crack  i\ 


If  Bnclc 
Wall,  i' 


Interior 
Surface.  - 


Coo*- at  each  End. 


f[                        Buulfc  -up  fcoof  \oq  on  3"  Concrete  tZoof  Deck. 

\ 

go 

m  

i    Lonqttudinoi  Axis  North  $  S 

1-  —  15  Windows  each 
side  +'-CfWidft 

M-P  

louth 

Lenqth  120'-  0* 
^ 

4 
1 

D< 

I 

«rt-Cra 

^-L 

BSJ             5*  Stonfc  Concr«fca  on 
fc>       y  Cinder  Concrete 
JT       onDtrt^ 

r 

Solid  U 

«xiDoo« 

ij 

FIG.  1.    ELEVATION  OF  FACTORY  BUILDING 


1.  LOCATION _ Philadelphia,  Pa. 

2.  LOWEST  OUTSIDE  TEMPERATURE.    (Table  2) —  6  F 

3.  BASE  TEMPERATURE:   In  this  example  a  design  temperature  10  F  above  lowest 

on  record  instead  of  15  F  is  used.    Hence  the  base  temperature  = 

(-  6  -f  10)  =  +  4  F. 

4.  DIRECTION  OF  PREVAILING  WIND  (during  Dec.,  Jan.,  Feb.) Northwest 

5.  BREATHING-LINE  TEMPERATURE  (5  ft  from  floor) 60  F 

6.  INSIDE  AIR  TEMPERATURE  AT  ROOF: 

The  air  temperature  just  below  roof  is  higher  than  at  the  breathing  line. 
Height  of  roof  is  16  ft,  or  it  is  16  —  5  =  11  ft  above  breathing  line.  Allowing 
2  per  cent  per  foot  above  5  ft,  or  2  X  11  =22  per  cent,  makes  the  tem- 
perature of  the  air  under  the  roof  =  1.22  X  60  =  73.2  F. 

7.  INSIDE  TEMPERATURE  AT  WALLS: 

The  air  temperature  at  the  mean  height  of  the  walls  is  greater  than  at 
the  breathing  line.  The  mean  height  of  the  walls  is  8  ft  and  allowing  2  per 
cent  per  foot  above  5  ft,  the  average  mean  temperature  of  the  walls  is 
1.06  X  60  =  63.6  F.  By  similar  assumptions  and  calculations,  the  mean 
temperature  of  the  glass  will  be  found  to  be  64.2  F  and  that  of  the  doors 
61.2  F. 

8.  AVERAGE  WIND  VELOCITY  (Table  2) 11.0  mph 

9.  OVER-ALL  DIMENSIONS  (See  Fig.  1) 120  x  50  x  16  ft 

10.  CONSTRUCTION: 

Walls — 12-in.  brick,  with  H-in.  plaster  applied  directly  to  inside  surface. 

Roof — 3-in.  stone  concrete  and  built-up  roofing. 

Floor — 5-in.  stone  concrete  on  3-in.  cinder  concrete  on  dirt. 

Doors— One  12  ft  x  12  ft  wood  door  (2  in.  thick)  at  each  end. 

Windows — Fifteen,  9  ft  x  4  ft  single  glass  double-hung  windows  on  each  side. 

11.  TRANSMISSION  COEFFICIENTS: 

Walls—  (Table 3,  Chapters,  WalI2B) U  =  0.34 

Roof—  (Table  11,  Chapter  5,  Roofs  2A  and  3A) U  =  0.77 

Floor—  (Table  10,  Chapter  5,  Floors  5A  and  6A) U  =  0.63 

Doors— (Table  13B,  Chapters) U  -  0.46 

Windows—  (Table  13A,  Chapters) U  =  1.13 

140 


CHAPTER  7 — HEATING  LOAD 


12.  INFILTRATION  COEFFICIENTS: 

Windows — Average  windows,  non-weatherstripped,  JlV"1-  crack  and 
%4-in.  clearance.  The  leakage  per  foot  of  crack  for  an  11-mile  wind 
velocity  is  25.0  cfh.  (Determined  by  interpolation  of  Table  2, 
Chapter  6.)  The  heat  equivalent  per  hour  per  degree  per  foot  of 
crack  is  taken  from  Chapter  6. 

25.0  X  0.018  =  0.45  Btu  per  deg  Fahr  per  foot  of  crack. 

Doors — Assume  infiltration  loss  through  door  crack  twice  that  of  windows 
or  2  X  0.45  =  0.90  Btu  per  deg  Fahr  per  foot  of  crack. 

Walls — As  shown  by  Table  1,  Chapter  6,  a  plastered  wall  allows  so  little 
infiltration  that  in  this  problem  it  may  be  neglected. 

13.  CALCULATIONS:     See  calculation  sheet,  Table  3. 


TABLE  3. 


CALCULATION  SHEET  SHOWING  METHOD  OF  ESTIMATING  HEAT  LOSSES  OF 
BUILDING  SHOWN  IN  FIG.  1 


PART  OF  BUILDING 

WIDTH 

IN 

FEET 

HEIGHT 

IN 

FEET 

NET  SUR- 
FACE AREA 
OR  CRACK 
LENGTH 

COEFFI- 
CIENT 

TEMP. 
DIFF. 

TOTAL 

BTU 

North  Wall: 
Brick,  H-in-  plaster  

50 
12 

1  pair 

16 
12 

doors 

656 
144 
60 

0.34 
0.46 
0.90 

59.6 
57.2 
57.2 

13,293 
3,789 
l,544a 

Doors  (2-in.  wood)  
\i  in.  Crack.__  

West  Wall: 
Brick,  H-in  plaster  

120 
15x4 
Double 
Window 

16 
9 
Hung 
re  (15) 

1380 
540 

450 

0.34 
1.13 

0.45 

59.6 
60.2 

60.2 

27,964 
36,734 

6,09  5a 

Glass  (Single) 

%  in.  Crack. 

South  Wall   _    

Same  as  North  Wall 

18,626 

East  Wall 

Same  as  West  Wall 

70.793 

Roof,  3-in.  concrete  and  slag- 
surfaced  built-up  roofing  

50 

120 

6000 

0.77 

69.2 

319,704 

Floor,  -^-in.  stone  concrete  on 
3-in.  cinder  concrete  „  

50 

120 

6000 

0.63 

5b 

18,900 

GRAND  TOTAL  of  heat  required  for  building  in  Btu  pei 

"  hour 

517,442 

"This  building  has  no  partitions  and  whatever  air  enters  through  the  cracks  on  the  windward  side  must 
leave  through  the  cracks  on  the  leeward  side.  Therefore,  only  one-half  of  the  total  crack  will  be  used  in 
computing  infiltration  for  each  side  and  each  end  of  building. 

bA  5  F  temperature  differential  is  commonly  assumed  to  exist  between  the  air  on  one  side  of  a  large 
floor  laid  on  the  ground  and  the  ground. 


PROBLEMS  IX  PRACTICE 

1  •  The  dry-bulb  temperature  and  the  relative  humidity  at  the  ceiling  of  a 
mixing  room  in  a  bakery  are  80  F  and  60  per  cent,  respectively.  The  roof  is  a 
4-in.  concrete  deck  covered  with  built-up  roofing.  If  the  lowest  outside  tem- 
perature to  be  expected  is  — 10  F,  what  thickness  of  rigid  fiber  insulation  will  be 
required  to  prevent  condensation? 

From  Table  11,  Chapter  5,  U  for  the  uninsulated  roof  =  0.72.  From  Table  2,  Chapter  5 , 
k  for  rigid  fiber  insulation  ==  0.33.  From  the  psychrometric  chart,  Chapter  1,  the  dew 
point  of  air  at  80  F  and  60  per  cent  relative  humidity  is  65  F.  The  ceiling  temperature, 
therefore,  must  not  drop  below  65  F  if  condensation  is  to  be  prevented. 

141 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

When  equilibrium  is  established,  the  amount  of  heat  flowing  through  any  component 
part  of  a  construction  is  the  same  for  each  square  foot  of  area. 

Therefore, 

^7  [80  -  (-10)]  -  1.65  (80  -  65) 
where 

U  is  the  transmittance  of  the  insulated  roof. 
Solving  the  equation,  U  «  0.275. 

The  resistance  of  the  insulated  roof  =    n  0?,    =3.64. 


The  resistance  of  the  uninsulated  roof  =    ,  70    =  1.39. 

U.  1  2i 

The  resistance  of  the  insulation  =  3.64  -  1.39  =  2.25. 
Resistance  per  inch  of  insulation  =   A  Q9    =3.0. 

U.Oo 

Since  a  resistance  of  2.25  is  required,  and  1  in.  of  insulation  has  a  resistance  of  3,  one  inch 
will  be  sufficient  to  prevent  condensation. 

The  same  result  might  have  been  obtained  by  selecting  an  insulated  4-in.  concrete  slab 
having  a  U  of  less  than  0.275  from  Table  11,  Chapter  5.  This  4-in.  concrete  slab  with 
1-in.  rigid  insulation  has  a  U  of  0.23  which  is  safe. 

2  •  What  inside  dry  -bulb  temperatures  are  usually  assumed  for:    (a)  homes, 
(b)  schools,  (c)  public  buildings? 

Referring  to  Table  1  : 

a.  70  to  72  F. 

b.  Temperature  varies  from  55  to  75  F,  depending  on  the  room.  Classrooms,  for  instance, 
are  usually  specified  as  70  to  72  F. 

c.  68  to  72  F. 

3  •  How  is  the  outside  temperature  selected  for  use  in  computing  heat  losses? 

The  outside  temperature  used  in  computing  heat  losses  is  generally  taken  from  10  to  15  F 
higher  than  the  lowest  recorded  temperature  as  reported  by  the  Weather  Bureau  during 
the  preceding  10  years  for  the  locality  in  which  the  heating  system  is  to  be  installed. 
In  some  cases  where  the  lowest  recorded  temperature  is  extremely  unusual,  the  design 
temperature  is  taken  even  higher  than  15  F  above  the  lowest  recorded  temperature. 

4  •  What  are  the  effects  of  wind  movement  on  the  heating  load? 

a.  Wind  movement  increases  the  heat  transmission  of  walls,  glass,  and  roof;  it  affects 
poor  walls  to  a  much  greater  extent  than  good  walls. 

b.  Wind  movement  materially  increases  the  infiltration  (inleakage)  of  cold  air  through 
the  cracks  around  doors  and  windows,  and  even  through  the  building  materials  them- 
selves if  such  materials  are  at  all  porous. 

5  •  Calculate  the  heat  given  off  by  eighteen  200-watt  lamps. 

200  X  18  X  3.415  =  12,294  Btu  per  hour. 

6  •  A  two-story,  six  room,  frame  house,  28-ft  by  30-ft  foundation,  has  the 
following  proportions: 

Area  of  outside  walls,  1992  sq  ft. 

Area  of  glass,  333  sq  ft. 

Area  of  outside  floors,  54  sq  ft. 

Cracks  around  windows,  440  ft. 

Cracks  around  doors,  54  ft. 

Area  of  second  floor  ceiling,  783  sq  ft. 

Volume,  first  and  second  floors,  13,010  cu  ft. 

Ceilings,  9  ft  high. 

142 


CHAPTER  7 — HEATING  LOAD 


The  minimum  temperature  for  the  heating  season  is  —  34  F,  and  the  required 
inside  temperature  at  the  30-in.  level  is  70  F.  The  average  number  of  degree 
days  for  a  heating  season  is  7851,  and  the  average  wind  velocity  is  10  mph, 
northwest. 

The  walls  are  constructed  of  2-in.  by  4-in.  studs  -with  wood  sheathing,  building 
paper,  and  wood  siding  on  the  outside,  and  wood  lath  and  plaster  on  the  inside. 
Windows  are  single  glass,  double-hung,  wood,  without  weatherstrips.  The 
second  floor  ceiling  is  metal  lath  and  plaster,  without  an  attic  floor.  The  roof 
is  of  wood  shingles  on  wood  strips  with  rafters  exposed.  The  area  of  the  roof  is 
20  per  cent  greater  than  the  area  of  the  ceiling.  Select  values  for  the  following: 
(a)  U  for  walls;  (b)  U  for  glass;  (c)  U  for  second  floor  ceiling;  (d)  U  for  roof; 
(e)  U  for  ceiling  and  roof  combined;  (f)  air  leakage,  cubic  feet  per  hour  per  foot 
of  window  crack;  (g)  air  leakage,  cubic  feet  per  hour  per  foot  of  door  crack. 

a.  0.25  (Table  5,  Chapter  5). 

b.  1.13  (Table  13,  Chapter  5). 

c.  0.69  (Table  8,  Chapter  5). 

d.  0.48  (Table  12,  Chapter  5). 

e.  0.236  (Equation  6,  Chapter  5). 
/.  21.4  (Table  2,  Chapter  6). 

g.  42.8,  which  is  double  the  window  leakage, 

7  •  Using  the  data  of  Question  6,  calculate  the  maximum  Btu  loss  per  hour  for 
the  various  constructions,  and  show  the  percentage  of  the  total  heat  which  is 
lost  through  each  construction  described. 

Assume  2  per  cent  rise  in  temperature  for  each  foot  in  height.  The  average  temperature 
will  be  72.8  F  for  walls,  doors,  and  windows,  and  79.1  F  for  the  second  floor  ceiling. 

a.  Outside  walls  46,200  Btu  loss  37.2  per  cent  of  total 

b.  Glass  34,950  Btu  loss  28.1  per  cent  of  total 

c.  Doors  5,670  Btu  loss  4.6  per  cent  of  total 

d.  Second  floor  ceiling  17,840  Btu  loss  14.3  per  cent  of  total 

e.  Air  leakage,  windows  15,750  Btu  loss  12.7  per  cent  of  total 
/.   Air  leakage,  doors  3,865  Btu  loss  3.1  per  cent  of  total 

Total  124,275  Btu  loss  100.0  per  cent  of  total 

8  •  For  the  house  in  Question  6,  place  1-in.  insulation  in  the  outside  walls  and 
second  floor  ceiling;  k  for  insulation  =  0.34.     Use  weatherstrip  on  doors  and 
windows,  and  double  glass  on  the  windows;  Ca  =  0.55.    Calculate  or  select  the 
following  values:  (a)  U  for  walls;  (b)  U  for  glass;  (c)  U  for  second  floor  ceiling; 
(d)  U  for  combination  of  ceiling  and  roof;  (e)  Air  leakage,  cubic  feet  per  hour 
per  foot  of  door  crack;  (f)  air  leakage,  cubic  feet  per  hour  per  foot  of  window 
crack. 

a.  0.144. 

b.  0.55. 

c.  0.23. 

d.  0.13. 

e.  15.5. 
/.   31.0. 

9  •  Calculate  the  maximum  Btu  loss  per  hour  and  show  the  percentage  loss  by 
each  channel  for  the  house  as  insulated  in  Question  8. 

a.  Outside  walls  26,650  Btu  loss  36.2  per  cent  of  total 

b.  Glass  17,000  Btu  loss  23.1  per  cent  of  total 

c.  Doors  5,670  Btu  loss  7.7  per  cent  of  total 

d.  Ceiling  10,070  Btu  loss  13.7  per  cent  of  total 

e.  Air  leakage,  windows  11,400  Btu  loss  15.5  per  cent  of  total 
/.   Air  leakage,  doors  2,795  Btu  loss  3.8  per  cent  of  total 

Total  73,585  Btu  loss        100.0  per  cent  of  total 

143 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

10  •  From  the  results  of  Questions  7  and  9,  calculate  the  Btu  saved  and  the 
percentage  saved  by  each  change  in  construction. 

Insulated  Uninsulated  Btu  Saved      Per  Cent  Saved 

a.  Outside  walls  46,200  26,650  19,550  42.3 

b.  Glass  34,950  17,000  17,950  51.4 

c.  Doors  5,670  5,670  0  0 

d.  Ceiling  3,865  2,795  1,070  27.7 

e.  Air  leakage,  windows  17,840  10,070  7,700  43.1 
/.  Air  leakage,  doors  15,750  11,400  4,350  27.6 

11  •  From  the  results  of  Questions  7  and  9,  calculate  the  heat  loads  per  heating 
season  in  Btu  and  note  the  savings  by  better  construction. 

The  7851  degree  days  for  the  heating  season  multiplied  by  24  hours,  times  the  Btu  loss 
per  hour  for  1  F  drop  in  temperature  gives  the  Btu  load  per  heating  season. 

Saving  =  250,800,000  -  148,000,000  =  102,800,000  Btu. 


Chapter  8 

COOLING  LOAD 

Conditions  to  be  Maintained,  Cooling  Load,  Transmission  for 
Surfaces  not  Exposed  to  the  Sun,  Outside  Temperatures,  Solar 
Radiation,  Time  Lag,  Transmission  of  Solar  Radiation  Through 
Glass,  Heat  and  Moisture  Leakage,  Heat  and  Moisture  Sources 

THE  method  of  calculating  the  cooling  load  is  similar  to  that  used 
in  calculating  the  heating  load.  The  direction  of  the  flow  of  heat  is 
reversed,  however,  and  in  most  cases  additional  factors  must  be  con- 
sidered, such  as  solar  radiation  and  the  heat  from  occupants,  lights, 
motors,  and  other  sources.  The  character  of  the  load  depends  on  the  type 
of  building  to  be  cooled  as,  for  example,  in  auditoriums  and  other  places 
of  assemblage  where  the  maximum  load  usually  is  that  due  to  the  heat  and 
moisture  given  off  by  the  occupants,  or  in  office  buildings  and  residences 
where  solar  radiation  and  the  transmission  and  infiltration  of  heat 
through  the  building  shell  are  most  important. 

While  cooling  is  generally  identified  with  the  summer  season,  it  is  often 
necessary  to  cool  in  winter  as  well  as  in  summer.  In  a  crowded  place  of 
assemblage  the  heat  given  off  by  the  occupants,  together  with  that  given 
off  by  the  lighting  and  power  equipment,  may  be  more  than  the  normal 
heat  loss  through  the  structure  even  in  winter  under  cold  climatic  con- 
ditions. 

Much  of  the  basic  information  for  the  design  of  comfort  conditioning 
installations  has  resulted  from  research  conducted  at  the  A.S.H.V.E. 
Research  Laboratory  and  at  institutions  with  which  cooperative  research 
investigations  have  been  carried  on.  These  data  include  the  effective 
temperature  index,  and  heat  and  moisture  loss  data  given  in  Chapter  2. 

COMFORT  CONDITIONS 

The  conditions  to  be  maintained  in  an  enclosure  are  variable  and 
depend  on  many  factors,  especially  the  season  of  the  year  and  (during  the 
summer)  the  outside  dry-bulb  temperature  and  the  duration  of  the  period 
of  occupancy.  Information  concerning  the  proper  effective  temperatures 
to  be  maintained  for  various  seasons  is  given  in  Chapter  2,  where  are  also 
tabulated  the  most  desirable  indoor  air  conditions  to  be  maintained  in 
summer  for  exposures  less  than  three  hours.  (See  Table  2,  Chapter  2.) 

In  installations  for  restaurants  and  theaters  the  requirements  are 
different  from  those  in  offices,  since  there  must  be  a  considerable  volume 
of  air  circulated  in  order  to  provide  ventilation  and  cooling. 

145 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


TABLE  1.    AVERAGE  MAXIMUM  DESIGN  DRY-BULB  TEMPERATURES,  DESIGN  WET-BULB 

TEMPERATURES,  WIND  VELOCITIES,  AND  WIND  DIRECTIONS  FOR 

JUNE,  JULY,  AUGUST,  AND  SEPTEMBER 


STATE 

CITY 

AVERAGE 
MAXIMUM 
DESIGN 
DRY-BULB 

DESIGN 
WET-BULB 

SUMMER  WIND 
VELOCITY 
MPH 

PREVAILING 
SUMMER  WIND 
DIRECTION 

Ala 

Birmingham.                 

93 

77 

5.2 

s 

Mobile.     

94 

78 

8.6 

sw 

Ariz 

Phoenix.          

HO 

77' 

6.0 

WT 

Ark 

Little  Rock. 

95 

77 

7.0 

NE 

Calif 

Los  Angeles    -         

88 

70 

6.0 

SW 

San  Francisco 

85 

68 

11.0 

sw 

Colo 

Denver  

90 

64 

6.8 

s 

Conn 

New  Haven 

88 

74 

7.3 

s 

D  C 

Washington  

95 

78 

6.2 

s 

Fla 

Jacksonville 

94 

78 

8.7 

sw 

Tampa.    _     

94 

79 

7.0 

E 

Ga. 

Atlanta  

91 

75 

7.3 

NW 

Savannah  

95 

79 

7.8 

SW 

Idaho 

Boise 

95 

65 

5.8 

NW 

111. 

Chicago    

95 

75 

10.2 

NE 

Peoria                            

91 

75 

8.2 

S 

Ind. 

Indianapolis    

90 

73 

9.0 

SW 

Iowa 

Des  Moines    - 

92 

74 

6.6 

sw 

Ky. 

Louisville..           

94 

75 

8.0 

sw 

La   

New  Orleans               

94 

79 

7.0 

sw 

Maine. 

Portland  

85 

71 

7.3 

s 

Md 

Baltimore                        

93 

76 

6.9 

sw 

Mass. 

Boston 

88 

73 

9.2 

sw 

Mich 

Detroit    .              

93 

73 

10.3 

sw 

Minn. 

Minneapolis  

84 

72 

8.4 

SE 

IVOss 

Vicksburg                 

95 

78 

6.2 

sw 

Mo. 

Kansas  City.  

92 

75 

9.5 

s 

St.  Louis  .                

95 

78 

9.4 

sw 

Mont. 

Helena  

87 

63 

7.3 

sw 

Nebr 

Lincoln  .  .               

93 

74 

9.3 

s 

Nev 

Reno 

93 

64 

7.4 

w 

N,  J. 

Trenton                       

95 

76 

10.0 

sw 

N.  Y.  .. 

Albany  

90 

74 

7.1 

s 

Buffalo.  ...              

83 

72 

12.2 

sw 

New  York.....  

95 

75 

12.9 

sw 

N.  M 

Santa  Fe..                    

87 

63 

6.5 

SE 

N.  C 

Asheville 

87 

72 

5.6 

SE 

Wilmington    

93 

79 

7.8 

sw 

N  Dak 

Bismarck. 

88 

69 

8.8 

NW 

Ohio 

Cleveland..    .  . 

95 

73 

9.9 

S 

Cincinnati.                          

95 

78 

6.6 

sw 

Okla. 

Oklahoma  City     

96 

76 

10.1 

s 

Oreg 

Portland 

83 

65 

6.6 

NW 

Pa. 

Philadelphia  

95 

78 

9.7 

SW 

Pittsburgh  

91 

73 

9.0 

NW 

R.  I. 

Providence     -.        

85 

73 

10.0 

NW 

S.  C. 

Charleston 

94 

80 

9.9 

SW 

Greenville.             

93 

76 

6.8 

NE 

Tenn 

Chattanooga 

94 

76 

6.5 

SW 

Memphis  ..         .     

93 

77 

7.5 

sw 

146 


CHAPTER  8 — COOLING  LOAD 


TABLE  1.    AVERAGE  MAXIMUM  DESIGN  DRY-BULB  TEMPERATURES,  DESIGN  WET-BULB 

TEMPERATURES,  WIND  VELOCITIES,  AND  WIND  DIRECTIONS  FOR 

JUNE,  JULY,  AUGUST,  AND  SEPTEMBER  (Continued) 


STATE 

CITY 

AVERAGE 
MASSMUM 
DESIGN 
DRY-BULB 

DESIGN 
WET-BULB 

SUMMER  WIND 
VELOCITY 
MPH 

PREVAILING 
SUMMER  WIND 
DIRECTION 

Texas 

Dallas 

99 

76 

94 

s 

Galveston  

93 

79 

97 

s 

San  Antonio 

100 

78 

74 

SE 

Houston  

93 

79 

7.7 

s 

El  Paso 

98 

69 

69 

E 

Utah 

Salt  Lake  City  

95 

67 

8.2 

SE 

Vt 

Burlington 

85 

71 

89 

s 

Va. 

Norfolk  

91 

76 

10.9 

s 

Richmond 

95 

78 

62 

SW 

Wash 

Seattle 

83 

61 

79 

s 

Spokane 

89 

63 

6  5 

SW 

WT  Va 

Parkersburg 

90 

74 

5  3 

SE 

Wis. 

Madison 

89 

73 

8  1 

SW 

Milwaukee         

93 

74 

10.4 

s 

Wyo. 
y        

Cheyenne 

85 

62 

92 

S 

COOLING  LOAD 

The  cooling  load  may  be  divided  into  the  following  parts: 

1.  Transmission  of  heat  through  walls,  roof,  and  glass  with   allowances  for  sun- 
exposed  surfaces  and  heat  capacity. 

2.  Transmission  of  solar  radiation  through  glass  and  absorption  by  interior  furnishings. 

3.  Heat  and  moisture  from  infiltration  and  from  outside  air  introduced. 

4.  Heat  and  moisture  from  occupants  and  heat  from  lights,  machinery  and  other 
sources. 

Transmission  for  Surfaces  Not  Exposed  to  the  Sun 

The  transmission  load  for  surfaces  not  exposed  to  the  sun  is  calculated  in 
a  manner  similar  to  that  described  in  Chapter  7,  by  means  of  the  following 
formula: 

Ht  =  AU(to-t)  (1) 

where 

Ht  =  heat  transmitted  through  the  material  of  the  wall,  glass,  roof,  or  floor,  Btu 

per  hour. 

A    =  net  inside  area  of  wall,  glass,  roof,  or  floor,  square  feet. 
t  —  inside  temperature,  degrees  Fahrenheit. 
to  =  outside  temperature,  degrees  Fahrenheit. 

U  —  coefficient  of  transmission  of  wall,  floor,  roof,  or  glass,  Btu  per  hour  per 
square  foot  per  degree  Fahrenheit  difference  in  temperature.  (Tables  3  to  13, 
Chapter  5.) 

Outside  Temperatures 

Summer  dry-bulb  and  wet-bulb  temperatures  for  various  -cities  are 
given  in  Table  1.  It  will  be  noted  that  the  temperatures  are  not  the 
maximums  but  the  design  temperatures  which  should  be  used  in  air- 
conditioning  calculations.  The  maximum  outside  wet-bulb  temperatures 

147 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


as  given  in  Weather  Bureau  reports  usually  occur  only  from  1  per  cent  to 
4  per  cent  of  the  time,  and  they  are  therefore  of  such  short  duration  that 
it  is  not  practical  to  design  a  cooling  system  covering  this  range.  The 
temperatures  shown  in  Table  1  have  been  chosen  after  extensive  study  of 
the  Weather  Bureau  records  and  are  temperatures  that  are  not  exceeded 
more  than  5  to  8  per  cent  of  the  time  during  June,  July,  August,  and 
September  for  an  average  year. 

Solar  Radiation 

Fig.  1  shows  the  total  amount  of  solar  energy  in  Btu  per  square  foot  per 
hour  received  during  the  day  by  a  surface  normal  to  the  rays  of  the  sun, 
by  a  horizontal  surface,  and  by  east,  west,  and  south  walls.  The  curves 
are  drawn  from  A.S.H.V.E.  Laboratory  data  obtained  by  pyrheliometer, 
are  based  on  sun  time,  and  are  for  a  perfectly  clear  day  on  August  1  at  a 
north  latitude  of  40  deg.  Data  from  these  curves  may  be  used  with 
little  error  for  most  United  States  latitudes  and  for  all  of  the  hotter 
months  of  the  year. 

The  absorption  of  solar  radiation  by  a  surface  depends  upon  the 
character  of  the  surface  and  the  angle  of  the  surface  with  respect  to  the 
direction  of  the  radiation.  The  heat  absorption  by  a  black  oilcloth 
surface  perpendicular  to  the  sun's  rays  was  found  to  be  as  high  as  273  Btu 
per  square  foot  per  hour,  based  on  tests  conducted  by  the  A.S.H.V.E. 
Research  Laboratory  in  Pittsburgh1.  Lamp  black,  red  brick  dust,  and 
aluminum  bronze  painted  surfaces  perpendicular  to  the  sun's  rays 
showed,  respectively,  94.0,  63.4,  and  28.2  per  cent  as  high  a  rate  of 
absorption  as  the  black  oilcloth. 

TABLE  2.    ALLOWANCE  FOR  SOLAR  RADIATION  ON  ROOFS  AND  WALLS 

APPROXIMATE  NUMBER  OF  DEGREES  TO  ADD  TO  DRY- BULB  TEMPERATURE 
FOR  DIFFERENT  TYPES  OF  SURFACES 


TYPE  OF  SURFACE 

BLACK 

RED  BRICK  OR  TILE 

ALTTMINUM  PA.  INT 

Roof  horizontal 

45 

30 

15 

East  or  west  wall                                       

30 

20 

10 

South  wall                                    -            

15 

10 

5 

Solar  radiation  is  an  important  factor  in  the  mechanism  of  heat  flow 
into  buildings.  Research  conducted  at  the  A.S.H.V.E.  Research  Labora- 
tory2 has  shown  that  a  large  error  may  be  introduced  into  the  calculations 
by  failure  to  consider  the  periodical  character  of  heat  flow  resulting  from 
the  diurnal  movement  of  the  sun  and  the  heat  capacity  of  the  structure, 
which  determine  the  timing  and  magnitude  of  the  heat  wave  flowing 
through  the  wall  into  a  building  on  a  hot,  sunny  day. 


Absorption  of  Solar  Radiation  in  Relation  to  the  Temperature,  Color,  Angle,  and  Other  Characteristics 
of  the  Absorbing  Surface,  by  F.  C.  Houghten  and  Carl  Gutberlet  (A.S.H.V.E.  TRANSACTIONS,  Vol.  36, 1930), 

2For  further  information  on  this  subject  see  following  A.S.H.V.E.  research  papers:  Coefficients  of  Heat 
Transfer  as  Measured  under  Natural  Weather  Conditions,  by  F.  C.  Houghten  and  C.  G.  F.  Zobel  (A.S.H. 
V.E.  TRANSACTIONS,  Vol.  34,  1928);  Absorption  of  Solar  Radiation  in  Its  Relation  to  the  Temperature, 
Color,  Angle  and  Other  Characteristics  of  the  Absorbing  Surface,  by  F.  C.  Houghten  and  Carl  Gutberlet 
(A.S.H.V.E.  TRANSACTIONS,  Vol.  36,  1930);  Heat  Transmission  as  Influenced  by  Heat  Capacity  and  Solar 
Radiation,  by  F.  C.  Houghten,  j.  L.  Blackshaw,  E.  M.  Pugh  and  Paul  McDermott  (A.S.H.V.E.  TRANS- 
ACTIONS, Vol.  38,  1932). 

148 


CHAPTER  8 — COOLING  LOAD 


Unfortunately,  the  calculations  for  the  transmission  of  heat  from  solar 
radiation  through  building  walls  are  too  complicated  to  be  of  much 
practical  value  to  the  heating  and  ventilating  engineer.  Approximate 
results  may  be  obtained  by  adding  the  number  of  degrees  given  in  Table  2 
to  the  outside  design  dry-bulb  temperature  in  calculating  the  heat  trans- 
mission through  a  wall  or  roof  which  may  be  exposed  to  the  sun  for  an 
appreciable  length  of  time.  Table  2  was  obtained  from  a  study  of  the 
data  in  A.S.H.V.E.  research  papers  on  solar  radiation1'  3.  Black  and 
aluminum  painted  surfaces  represent  the  extremes  which  are  likely  to 
occur.  For  other  types  of  surfaces,  values  intermediate  between  those 
given  in  the  table  can  be  used. 

Time  Lag 

The  calculation  of  heat  transmitted  through  walls  and  roofs  does  not 
take  into  consideration  the  heat  capacity  of  the  structure  and  the  con- 
sequent time  lag  in  the  transmission  of  heat.  In  the  thick  walls  used  in 
modern  office  buildings  the  time  lag  may  amount  to  10  hours  or  more4. 
Thus  in  many  cases  the  wall  transmission  cannot  be  added  directly  to  the 
cooling  load  from  other  sources  because  the  peak  of  the  wall  transmission 
load  may  not  coincide  with  the  peak  of  the  total  cooling  load  and  may 
even  occur  after  the  cooling  system  has  been  shut  down  for  the  day.  The 
data  in  Table  3  were  taken  from  A.S.H.V.E.  research  papers3'  4  and 
while  they  result  principally  from  a  study  of  experimental  slabs,  they  give 
an  idea  of  the  time  lag  to  be  expected  in  various  structures. 

TABLE  3.  TIME  LAG  IN  TRANSMISSION  OF  SOLAR  RADIATION  THROUGH  WALLS  AND  ROOFS 


TYPE  AND  THICKNESS  OP  WALL  OR  ROOF 


TIME  LAG, 
HOTTRS 


2-in.  pine - _ 

6-in.  concrete 

4-in.  gypsum 

3-in.  concrete  and  1-in.  cork.. 

2-in.  iron  and  cork  (equivalent  to  %-in.  concrete  and  2.15-in.  cork)... 
4-in.  iron  and  cork  (equivalent  to  5j^-in.  concrete  and  1.94-in.  cork).. 
8-in.  iron  and  cork  (equivalent  to  16-in.  concrete  and  1.53-in.  cork).. 


19 


22-in.  brick  and  tile  wall _._.j     10 

In  intermittently  cooled  buildings  an  excess  cooling  capacity  must  be 
provided  to  care  for  the  additional  load  imposed  by  the  necessity  to  cool 
down  the  furnishings  and  the  material  of  the  interior  construction  to  the 
point  of  maintained  temperatures. 

Transmission  of  Solar  Radiation  Through  Glass 

In  considering  the  transmission  through  glass  several  factors  must  be 
considered.  As  the  sun's  rays  impinge  against  a  pane  of  glass,  most  of  the 
radiation  passes  through  to  the  other  side,  a  small  amount  is  reflected,  and 
the  balance  is  absorbed  by  the  glass.  The  amount  absorbed  depends  upon 


'Heat  Transmission  as  Influenced  by  Heat  Capacity  and  Solar  Radiation,  by  F.  C.  Houghten,  J.  L. 
Blacksnaw,  E.  M.  Pugh,  and  Paul  McDermott  (A.S.H.V.E.  TRANSACTIONS,  Vol.  38,  1932). 

'Field  Studies  of  Office  Building  Cooling  (A.S.H.V.E.  Research  Paper),  by  J.  H.  Walker,  S.  S.  Sanford, 
and  E.  P.  Wells  (A.S.H.V.E.  TRANSACTIONS,  Vol.  38,  1932). 

149 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  193-5 


rue 


CHAPTER  8 — COOLING  LOAD 


the  character  and  thickness  of  the  glass  and  the  angle  between  the  rays  of 
sunlight  and  the  glass.  The  temperature  of  the  glass  is  raised  by  the 
absorbed  heat  and  this  heat  is  then  delivered  to  the  air  on  the  two  sides  of 
the  glass  in  proportion  to  the  difference  between  glass  and  air  tem- 
peratures. 

The  A.S.H.V.E.  tests  indicated  that  a  single  pane  of  double  strength 
glass  0.127  in.  thick  absorbs  approximately  11  per  cent  of  the  solar 
radiation  passing  through  it  when  the  impingement  is  normal.  For 
smaller  angles  of  impingement,  the  glass  retards  percentages  of  the  total 
radiant  energy  approximately  in  proportion  to  the  sine  of  the  angle. 
Other  experiments4  indicate  a  glass  absorption  of  16.7  per  cent  for  one 
pane  of  glass  and  37.5  per  cent  for  two  J^-in.  panes  separated  by  a  1%-in. 
air  space. 

The  amount  of  solar  radiation  delivered  to  an  unshaded  glass  surface 
may  be  obtained  from  the  curves  in  Fig.  1.  For  surfaces  other  than  those 
given,  the  solar  radiation  incident  to  the  glass  must  be  calculated. 
Hendrickson  and  Walker6  have  shown  how  this  may  be  done  if  the  wall 
faces  some  direction  other  than  east,  west,  or  south.  They  have  also 
shown  how  to  calculate  the  net  glass  area  on  which  the  solar  radiation 
impinges  when  the  glass  is  partly  shaded  by  the  frame  or  wall.  The 
values  from  Fig.  1  must  be  used  only  for  the  net  glass  area  on  which  the 
sun  shines.  Recent  tests  at  the  A.S.H.V.E.  Research  Laboratory6  have 
determined  the  percentage  of  heat  from  solar  radiation  actually  delivered 
to  a  room  with  bare  windows  and  with  various  types  of  outdoor  and 
indoor  shading.  The  data  in  Table  4  are  taken  from  these  tests. 

TABLE  4.    SOLAR  RADIATION  TRANSMITTED  THROUGH  BARE  AND  SHADED  WINDOWS 


PER  CENT  DELIVERED 
TO  ROOM 

Bare  window  glass  

97 

Canvas  awning                  .          .... 

28 

Inside  shade,"  fully  drawn 

45 

Inside  shade,  one-half  drawn 

68 

Inside  Venetian  blind,  fully  covering  window 

58 

Outside  Venetian  blind,  fully  covering  window  

22 

The  percentage  figures  in  this  table  were  obtained  by  dividing  the  total 
amount  of  heat  actually  entering  through  the  shaded  window  by  the 
total  amount  of  heat  calculated  to  enter  through  a  bare  window  (solar 
radiation  plus  glass  transmission  based  on  observed  outside  glass  tem- 
perature). For  bare  windows  on  which  the  sun  shines,  the  transmission 
of  heat  from  outside  air  to  glass  is  small  as  the  glass  temperature  is  raised 
by  the  solar  radiation  absorbed.  Therefore,  in  calculating  the  total  heat 
gain  through  windows  on  the  sunny  sides  of  buildings,  it  is  sufficiently 
accurate  to  figure  the  total  cooling  load  due  to  the  window,  as  the  solar 
radiation  times  the  proper  factor  from  Table  4,  and  to  neglect  the  heat 


*Summer  Cooling  for  Comfort  as  Affected  by  Solar  Radiation,  by  G.  A.  Hendrickson  and  ].  H.  Walker, 
Heating  and  Ventilating,  November,  1932,  and  The  Determination  of  Sun  Effect  on  Summer  Cooling  Loads, 
by  G.  A.  Hendrickson  and  J.  H.  Walker,  Heating  and  Ventilating,  June,  1933. 

^Studies  of  Solar  Radiation  Through  Bare  and  Shaded  Windows,  by  F.  C.  Houghten,  Carl  Gutberlet, 
and  J.  L.  Blackshaw  (A.S.H.V.E,  Journal  Section,  Heating,  Piping  and  Air  Conditioning,  February,  1934). 

151 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

transmission  through  the  glass  caused  by  the  difference  between  the 
temperatures  of  the  inside  and  outside  air.  Another  reason  for  neglecting 
this  glass  transmission  load  is  that  the  curves  in  Fig.  1  were  based  on  the 
maximum  intensity  of  solar  radiation  observed  at  the  A.S.H.V.E.  Labora- 
tory during  a  three-year  study,  so  results  based  on  these  curves  will  be 
amply  high.  It  will  be  noted  that  Table  4  gives  the  amount  of  heat 
delivered  through  the  window  as  97  per  cent  of  the  solar  radiation,  which 
is  greater  than  is  indicated  by  the  figures  for  absorption  in  the  preceding 
paragraph.  The  explanation  is  that  much  of  the  radiation  absorbed  by 
the  glass  is  delivered  to  the  room. 

Fig.  1  shows  that  the  maximum  solar  intensity  on  any  surface  is  of 
limited  duration.  In  the  case  of  windows  the  total  energy  impinging  on 
the  glass  before  and  after  the  time  of  maximum  intensity  is  further 
reduced  by  increased  shading  of  the  glass  from  the  frame,  or  wall.  The 
cooling  load  due  to  solar  radiation  therefore  does  not  have  to  be  figured  as 
a  steady  load.  Another  point  which  should  be  noted  is  that  the  maximum 
solar  radiation  load  on  an  east  wall  occurs  early  in  the  morning  when  the 
outside  temperature  is  low. 

In  a  recent  paper  by  the  A.S.H.V.E.  Research  Laboratory7  it  was  shown 
that  ordinary  double  strength  window  glass  transmits  no  measurable 
amount  of  energy  radiated  from  a  source  at  500  F  or  lower ;  that  it  trans- 
mits only  6.0  and  12.3  per  cent  of  the  total  radiation  from  surfaces  at 
700  F  and  1000  F,  respectively;  and  that  it  transmits  65.7  per  cent  of  the 
radiation  from  an  arc  lamp,  76.3  per  cent  of  the  radiation  from  an  in- 
candescent tungsten  lamp,  and  89.9  per  cent  of  the  radiation  from  the 
sun.  Thus,  glass  windows  in  a  room  constitute  heat  traps,  which  allow 
rather  free  transmission  of  radiant  energy  into  the  room  from  the  sun  to 
warm  objects  in  it,  but  do  not  allow  the  transmission  of  re-radiated  heat 
from  these  same  objects. 

Some  recent  tests4  indicated  that  sunshine  through  window  glass  is 
the  most  important  factor  to  contend  with  in  the  cooling  of  an  office 
building.  At  times  it  was  shown  to  account  for  as  much  as  75  per  cent  of 
the  total  cooling  necessary.  Because  of  the  importance  of  the  sunshine 
load,  cooling  systems  should  be  zoned  so  that  the  side  of  the  building  on 
which  the  sun  is  shining  can  be  controlled  separately  from  the  other  sides 
of  the  building.  If  buildings  are  provided  with  awnings  so  that  the 
window  glass  is  shielded  from  sunshine,  the  amount  of  cooling  required 
will  be  reduced  and  there  will  also  be  less  difference  in  the  cooling  require- 
ments of  different  sides  of  the  building.  The  total  cooling  load  for  a 
building  exposed  to  the  sun  on  more  than  one  side  is  of  course  less  than 
the  sum  of  the  maximum  cooling  loads  in  the  individual  rooms  since  the 
maximum  solar  radiation  load  on  the  different  sides  occurs  at  different 
times. 

Heat  and  Moisture  Leakage 

An  allowance  must  be  made  for  the  heat  and  moisture  in  the  outside  air 
introduced  for  ventilating  purposes  or  entering  the  building  through 
cracks,  crevices,  doors,  and  other  places  where  infiltration  might  occur. 


'Radiation  of  Energy  Through  Glass,  by  J.  L.  Blackshaw  and  F.  C.  Houghten  (A.S.H.V.E.  Journal 
Section,  Heating,  Piping  and  Air  Conditioning,  October,  1933). 

152 


CHAPTER  8 — COOLING  LOAD 


The  volume  of  air  entering  due  to  infiltration  may  be  estimated  from  data 
given  in  Chapter  6,  and  information  on  the  amount  of  outside  air  required 
for  ventilation  will  be  found  in  Chapter  2. 

The  heat  gain  resulting  from  the  outside  air  introduced  may  be  esti- 
mated from  the  following  formula: 

Hi  =  Qd0  (00  -  0)  (2) 

where 

Hi  =  heat  to  be  removed  from  outside  air  entering  the  building,  Btu  per  hour. 
Q  —  volume  of  outside  air  entering  the  building,  cubic  feet  per  hour. 
d0  =  density  of  outside  air,  pounds  of  dry  air  per  cubic  foot  of  outside  air,  at  the 

temperature  A> 
©o  ~  heat  content  of  mixture  of  outside  dry  air  (at  temperature  to)  and  water  vapor, 

Btu  per  pound  of  dry  air. 

©  =  heat  content  of  mixture  of  inside  dry  air  (at  temperature  /)  and  water  vapor, 
Btu  per  pound  of  dry  air. 

Heat  and  Moisture  Sources 

Figs.  6  to  9,  Chapter  2,  show  the  heat  and  moisture  given  off  by  human 
beings  under  various  conditions  of  activity.  For  average  conditions  where 
a  person  is  normally  at  rest,  as  in  a  theater,  or  doing  very  light  work,  as  in 
a  restaurant  or  residence,  the  total  amount  of  heat  given  off  will  average 
about  400  Btu  per  hour.  Part  of  this  is  latent  heat  due  to  the  evaporation 
of  700  to  1200  grains  of  moisture  per  hour.  Examples  illustrating  heat  and 
moisture  loss  calculations  for  human  beings  are  given  in  Chapter  2. 

TABLE  5.    HEAT  GAIN  DUE  TO  VARIOUS  DEVICES,  BTU  PER  HOUR 


Lights  and  electric  appliances  

3,415  per  kilowatt 

Motors,  X-JLO  hp 

255 

Motors,  1  hp  

2,546 

Restaurant  coffee  urns,  10-gal  capacity  

16",000 

Dish  warmers  per  10  sq  ft  of  shelf 

6,000 

Restaurant  range  —  4  burners  and  oven  

100,000 

Residence  gas  range 
Giant  burner  

12,000 

Medium  burner 

9,000 

Oven  

1,000  per  cu  ft  of  space 

Pilot..     .. 

250 

Electric  Range 
Small  burner,  100  to  1350  watts              

3,415  to  4,600 

Large  burner,  1700  to  2200  watts  

5,800  to  7,500 

Oven,  2000  to  3000  watts 

6,830  to  10,245 

Appliance  connection,  660  watts  

2,250 

Warming  compartment,  300  watts  

1.025 

All  sources  of  heat  must  of  course  be  considered  in  designing  the  con- 
ditioning system.  The  heat  gain  due  to  various  devices  is  given  in 
Table  5.  An  example  of  cooling  load  calculation  is  given  in  Chapter  9. 

PROBLEMS  IN  PRACTICE 

1  •  a.  What  should  be  the  dry-  and  wet-bulb  temperatures  in  a  restaurant 
when  the  outdoor  dry-bulb  temperature  is  95  F? 

b.  "What  is  the  most  desirable  indoor  dry -bulb  temperature  and  relative 
humidity  in  an  office  building  in  summer? 

a.  Dry-bulb,  80  F;  wet-bulb,  65  F.     (Table  2,  Chapter  2.) 

b.  76.5  F  and  50  per  cent  relative  humidity.     (Fig.  3,  Chapter  2.) 

153 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

2  •  The  outdoor  and  indoor  temperatures  are  90  F  and  78  F,  respectively. 
What  is  the  amount  of  heat  transmitted  per  hour  through  a  7  ft  by  4  ft  north 
window? 

Ht  =  28  X  1.13  (90-  78)  =  380  Btu  per  hour. 
(Equation  1,  Chapter  8  and  Table  12,  Chapter  5.) 

3  •  What  are  the  proper  design  temperatures  for  a  Detroit  store? 

Outdoor  dry-bulb,  88  F;  wet-bulb,  72  F.    (Table  1,  Chapter  8.) 
Indoor  dry-bulb,  77.5  F;  wet-bulb,  64.5  F.    (Table  2,  Chapter  2.) 

4  •  a.  What  is  the  maximum  heat  transmission  for  a  flat  roof  exposed  to  the 
sun  with  the  outdoor  and  indoor  temperature  95  F  and  80  F,  respectively?    The 
roof  is  of  uninsulated  6-in.  concrete,  with  its  underside  exposed,  and  with  a 
black  upper  surface. 

b.  If  the  temperatures  specified  were  the  maximum  for  the  day  and  occured 
at  12  o'clock,  at  what  time  would  the  maximum  cooling  load  due  to  the  roof 
exist? 

a.  Ht  =  1  X  0.64  (95  +  45  -  80)  »  38.4  Btu  per  hour  per  square  foot. 
(Equation  1  and  Table  2,  Chapter  8,  and  Table  11,  Chapter  5.) 

b.  At  3  p.m.    (Table  3.) 

5  •  For  south  windows  equipped  with  canvas  awnings,  what  is  the  maximum 
amount  of  heat  delivered  to  a  room  when  the  outdoor  temperature  is  90  F  and 
the  indoor  temperature  is  78  F? 

115  X  0.28  =  32.2  Btu  per  square  foot  of  glass  (Fig.  1  and  Table  4;  note  that  glass 
transmission  can  be  neglected). 

6  •  What  is  the  heat  gain  per  cubic  foot  of  outside  air  introduced,  under  the 
following  conditions : 

Outdoor  temperatures,  90  F  dry-bulb  and  75  F  wet-bulb. 
Inside  temperatures,  78  F  dry-bulb  and  65  F  wet -bulb. 

Hi  =  Qdo  (©o  -  ©).    Equation  2. 

The  relative  humidity  of  the  outdoor  air  is  50  per  cent  (Fig.  3,  Chapter  2),  and  d0  — 
I 


14.21 


=  0.0703  (Table  5,  Chapter  1). 


©  «  37.81  and  ©  =  29.65  (Table  5,  Chapter  1).  The  total  heat  of  any  air- vapor  mix- 
ture may  be  obtained  from  the  last  column  in  Table  5,  Chapter  1,  by  considering  the 
temperatures  to  be  wet-bulb  readings,  since  the  total  heat  of  a  mixture  is  constant  for  a 
given  wet-bulb  temperature. 

Hi  =  1  X  0.0703  (37.81  -  29.65)  =  0.57  Btu  per  cu  ft. 

7  •  If  there  are  twenty  200 -watt  lights  in  use  in  a  room,  what  is  the  cooling 
load  due  to  lights? 

200  X  20  »  4000  watts  =  4  kw. 

3415  X  4  =  13,660  Btu  per  hour  (Table  5,  Chapter  8). 

8  •  a.  When  a  restaurant  has  two  10-gal  coffee  urns,  what  is  the  cooling  load 
due  to  them? 

b.  What  is  the  cooling  load  due  to  four  1350 -watt  burners  on  an  elecjtric 
range? 

a.  16,000  X  2  «  32,000  Btu  per  hour  (Table  5,  Chapter  8). 

b.  4600  X  4  =  18,400  Btu  per  hour  (Table  5,  Chapter  8). 

154 


Chapter  9 

CENTRAL  AIR  CONDITIONING 
SYSTEMS 

Types  of  Systems,  Dehumidifier  s,  Designing  the  System,  Zoning, 
Location  of  Apparatus,  Temperature  of  the  Air  Leaving  Outlets, 
Air  Quantity  Required,  Heat  to  be  Removed  by  Cooling  and 
Dehumidifying  Apparatus,  Size  of  Reheaters,  Surface  Cooling 
Problems,  Auxiliary  Equipment 


systems,  equipped  for  cooling  and  dehumidifying,  are  used 
_  principally  in  the  air  conditioning  of  theaters,  restaurants,  office 
buildings,  or  other  places  where  many  people  gather,  and  in  manufacturing 
establishments  where  air  conditions  have  an  important  influence  on  the 
quality  of  product  or  rate  of  production.  A  central  cooling  and  de- 
humidifying  plant  is  one  in  which  the  fans,  dehumidifiers,  and  other 
related  apparatus  are  assembled  in  suitable  apparatus  rooms  from  which 
distribution  and  return  ducts  lead  to  the  conditioned  spaces.  The  design 
of  such  systems  is  considered  in  this  chapter,  while  in  Chapter  22  central 
systems  for  heating  and  humidifying  are  described.  Industrial  air  con- 
ditioning has  been  considered  in  Chapter  3. 

TYPES  OF  SYSTEMS 

Dehumidification  or  cooling  of  air  may  be  accomplished  by  several 
methods  and  by  use  of  many  heat  transfer  mediums.  Most  comfort- 
conditioning,  central  station,  air-conditioning  systems  employ  cold  water 
or  the  direct  expansion  of  a  refrigerant  in  either  spray  type  or  surface 
type  equipment  to  accomplish  the  required  cooling  and  dehumidification. 
Among  the  several  other  methods  that  may  be  employed  are  :  passing  the 
air  through  or  over  a  dehydrating  agent  and  then  lowering  the  dry-bulb 
temperature  to  the  proper  level,  and  evaporative  cooling.  The  former 
method  is  applicable  to  comfort  conditioning  only  where  reasonably  cold 
water  is  available  for  reducing  the  dry-bulb  temperature  after  dehydra- 
tion, while  the  latter  method  is  applicable  to  comfort  conditioning  only  in 
regions  where  the  summer  wet-bulb  temperature  is  low. 

If  the  system  is  intended  solely  for  summer  conditioning,  the  apparatus 
will  consist  essentially  of  a  dehumidifier  of  the  surface  type  or  spray  type  ; 
filters;  fan  and  motor;  reheater;  outside  air,  return  air,  and  supply  air  duct 
work;  air  outlets  and  grilles;  spray  pump  for  spray  dehumidifier;  refrigera- 
tion equipment;  and  suitable  controls.  Generally,  however,  a  central 
station  air  conditioning  system  is  designed  for  year-round  service.  This 
means  that  properly  sized  heaters  and  humidifiers,  with  their  respective 

155 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

controls,  must  be  added.  With  few  exceptions,  systems  designed  to  meet 
summer  capacity  requirements  will  have  ample  capacity  for  winter  and 
intermediate  season  conditioning. 

A  common  arrangement  of  a  central  station  spray  type  system  for 
cooling  and  dehumidifying  is  illustrated  in  Fig.  1.  The  plant  may  be 
designed  to  condition  100  per  cent  outside  air,  100  per  cent  return  air,'or  a 
mixture  of  outside  and  return  air.  Further,  part  of  the  air  returned  from 
the  conditioned  space  may  be  by-passed1  around  the  conditioner  as 
illustrated  in  Fig.  2.  The  reheater  may  be  installed  in  the  fan  inlet 
chamber  as  shown,  in  the  by-pass  air  duct,  or  in  the  fan  discharge  duct, 
depending  upon  apparatus  space  and  other  design  conditions.  Still 
another  arrangement  of  equipment  will  result  if  the  dehumidified  air  fan 
delivers  the  conditioned  air  to  several  other  fans  rather  than  to  the  con- 


Outside 


FIG.  1.    SPRAY  TYPE  AIR  CONDITIONING  APPARATUS 


ditioned  space  directly.  These  booster  fan  equipments  may  use  part  by- 
pass air  as  illustrated  in  Fig.  3  or  100  per  cent  dehumidified  air  and 
reheaters.  The  main  apparatus,  in  either  case,  may  or  may  not  have  a 
by-pass  connection,  depending  on  load  conditions  and  other  design  factors. 
The  systems  illustrated  in  Figs.  1  and  2  may  be  converted  into  the 
surface  cooling  type  by  merely  replacing  the  dehumidifiers  with  surface 
cooling  coils  which  use  cold  water  or  direct  expansion  of  refrigerant  to 
accomplish  the  required  cooling  and  dehumidifying.  The  coils  may  also 
be  installed  within  the  spray  chamber,  either  in  series  with  the  sprays 
or  below  them. 

DEHUMIDIFIERS 

Information  on  spray  type  dehumidifiers  is  given  in  Chapter  11. 

Surface  cooling  type  dehumidifiers  generally  consist  of  extended-surface 
coils  within  which  the  water  or  refrigerant  is  circulated  or  the  refrigerant 
is  expanded.  The  air  to  be  cooled  and  dehumidified  is  drawn  or  blown 
over  the  coils.  This  system  is  generally  comparatively  low  in  initial  cost 
and  has  low  operating  costs.  For  comfort  cooling,  water  is  usually  used  to 


Patents  exist  covering  the  use  of  the  by-pass  for  cooling  and  dehumidifying  systems. 

156 


CHAPTER  9 — CENTRAL  AIR  CONDITIONING  SYSTEMS 


bring  the  refrigeration  effect  to  the  coils.  Many  localities  have  refrigera- 
tion codes  which  restrict  the  use,  in  comfort  conditioning  applications,  of 
refrigerants  acting  by  direct  expansion  in  coils  exposed  to  the  air  stream. 
Therefore,  local  codes  should  be  consulted  by  the  designer  before  he  plans 
a  system  employing  direct-expansion  methods.  Close  humidity  control 
cannot  be  maintained  during  the  cooling  season  by  the  surface  cooling 
type  of  equipment.  Winter  humidification  may  be  accomplished  by  use 
of  evaporating  pans  or  spray  nozzles.  The  cooling  coils  serve  no  purpose 
during  the  intermediate  or  heating  seasons,  so  in  this  respect  the  spray 
type  equipment  is  often  preferred,  in  that  during  certain  seasons  evapora- 
tive cooling  will  be  sufficient  to  produce  the  cooling  desired.  Effective 
cooling  and  dehumidification  accomplished  by  surface  units  are  dependent 
upon  many  variable  factors.  The  air  velocity  through  the  unit,  air 


FIG.  2.    SPRAY  TYPE  AIR  CONDITIONING  APPARATUS  WITH  BY- PASS 


temperature,  moisture  content  of  the  air,  water  or  refrigerant  tempera- 
ture, and  velocity  of  the  water  or  refrigerant  through  the  tubes  must  be 
considered  in  selecting  the  proper  unit  for  a  given  design  load.  If  any  of 
these  factors  vary  without  a  corresponding  variation  of  the  other  factors, 
the  effective  work  of  the  coil  will  increase  or  decrease,  as  the  case  may  be. 

DESIGNING  THE  SYSTEM 

The  general  procedure  for  the  design  of  a  central  cooling  and  de- 
hum  Jdifying  system  is  as  follows  : 

1.  Calculate  the  heat  gain  for  each  room  or  space  to  be  conditioned.    (See  Chapters 
5  and  8.) 

2.  Determine  the  volume  of  outside  air  to  be  introduced.    (See  Chapter  2.) 

3.  Assume  or  calculate  the  temperature  of  air  leaving  the  supply  outlets. 
Calculate  the  quantity  of  air  to  be  circulated. 

Estimate  the  temperature  loss  in  the  duct  system. 

Calculate  the  heat  to  be  removed  by  the  cooling  and  dehumidifying  apparatus. 

Calculate  the  size  of  the  reheating  equipment. 

8.  Select  cooling  equipment  and  heating  equipment  from  manufacturers'  data  and 
performance  curves. 

9.  Calculate  total  tonnage. 

157 


4. 
5. 
6. 
7. 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


10.  Design  the  air  distribution  system  and  the  air  outlets  and  inlets.     (See  Chapters 
19  and  20.) 

11.  Calculate  the  total  static  pressure  of  the  system. 

12.  Select  the  fan,  motor,  and  drive.    (See  Chapter  17.) 

13.  Select  the  pump  and  motor. 

14.  Design  the  control  system.    (See  Chapter  14.) 

ZONING 

The  above  general  outline  of  procedure  will  prove  satisfactory  for  the 
smaller  and  less  complex  installations.  However,  when  dealing  with  air- 
conditioning  systems  for  large  buildings,  after  a  proper  analysis  has  been 
made  of  the  conditions  to  be  maintained  and  the  heat  loads  encountered, 
it  is  generally  considered  best  practice  to  divide  the  complete  job  into  a 

Motor 

t° 


t 


s 


ggg       , 

|Fanl      |        I  To  room  B  • 


FIG.  3.    CENTRAL  DEHUMIDIFYING  PLANT  AND  LOCAL  RECIRCULATING  FANS 

number  of  suitably  sized  units.  In  some  cases  a  unit  per  floor  or  group  of 
floors  may  complete  the  design  satisfactorily,  whereas  in  others  it  may  be 
advantageous  to  have  separate  units  for  each  of  the  various  outside 
exposures  of  the  building.  Where  the  floor  area  is  large  in  relation  to  the 
outside  wall  exposure,  it  is  obvious  that  provision  must  be  made  for  the 
variable  load  to  which  the  outside  exposures  are  subjected.  The  heat 
loads  on  inside  rooms  are  apt  to  be  less  variable  since  the  fluctuations  of 
the  outside  weather  conditions  are  not  directly  involved.  Such  conditions 
often  result  in  the  natural  zoning  or  segregation  of  rooms  having  similar 
exposures  and  internal  heat  loads. 

LOCATION  OF  APPARATUS 

Availability  of  space  for  apparatus  and  duct  work  is  of  primary  im- 
portance when  selecting  the  type  of  system  for  a  given  design.  In  general, 
for  large  installations,  the  refrigeration  equipment,  because  of  its  size, 

158 


CHAPTER  9 — CENTRAL  AIR  CONDITIONING  SYSTEMS 


weight,  and  operating  characteristics,  is  located  in  the  basement  along 
with  the  boilers,  fire  pumps,  and  other  equipment.  The  air  conditioning 
apparatus  is  generally  located  where  clean  outdoor  air  is  readily  available, 
the  designer  bearing  in  mind  that  supply  and  return  air  ducts,  steam  con- 
nections, water  and  drain  connections,  and  electrical  connections  must  be 
made  to  the  equipment  proper. 

TEMPERATURE  OF  AIR  LEAVING  OUTLETS 

In  comfort  conditioning  applications,  air  has  been  distributed  from 
properly  designed  outlets  without  producing  drafts  at  temperatures 
varying  from  approximately  five  to  thirty  degrees  below  the  required 
room  temperature.  Factors  influencing  the  design  and  selection  of  air 
outlets  are:  ceiling  height,  type  of  ceiling,  length  of  blow,  and  temperature 
and  quantity  of  air  to  be  distributed.  Most  summer  conditioning  instal- 
lations are  designed  to  supply  the  air  to  the  conditioned  space  at  from 
8  to  15  deg  below  room  temperature.  Recently  the  use  of  specially 
designed  nozzles  has  indicated  the  possibility  of  reducing  the  air  quantity 
necessary  to  dissipate  a  given  heat  load  by  introducing  the  air  into  the 
room  as  much  as  thirty  degrees  below  room  temperature.  Comfort  con- 
ditioning systems  employing  differentials  greater  than  fifteen  degrees 
require  special  consideration  and  design  experience  because  high  pressure 
outlets  or  nozzles  are  usually  used.  Further,  care  mustte  taken  to  allow 
a  sufficient  air  quantity  under  all  load  conditions  to  insure  good  distri-r 
bution.  If  winter  heating,  as  well  as  summer  conditioning,  is  to  be  accom- 
plished by  the  same  distributing  system,  the  design  of  the  outlets  will  be 
influenced  as  discussed  in  Chapter  22.  Industrial  systems  in  which  drafts 
are  not  objectionable  usually  employ  a  temperature  differential  equal  to 
the  dew-point  depression. 

AIR  QUANTITY  REQUIRED 

For  calculating  the  quantity  of  air  required  to  absorb  a  given  heat  gain, 
the  following  approximate  formulae  may  be  used  : 

M  =  * 


60  X  0.24  X  (t  -  t 

or,  assuming  a  constant  value  of  0.075  Ib  for  d, 

_     g.  X  55.2 


~  60  X  (t  -  ty) 
where 

Q  =  volume  of  air  required,  cubic  feet  per  minute. 
Hs  =  total  sensible  heat  gain,  Btu  per  hour. 
/  =  room  temperature,  degrees  Fahrenheit. 
ty  =  outlet  temperature,  degrees  Fahrenheit. 
M  =  weight  of  air  required,  pounds  per  minute. 

d  —  density  of  air  at  the  temperature  and  relative  humidity  of  the,  room,  pounds  per 
cubic  foot. 

Example  1.  The  total  sensible  heat  gain  in  a  restaurant  when  held  at  80  F  is  190,736 
Btu  per  hour.  Assuming  a  12  deg  Fahr  temperature  differential  between  the  entering 
air  and  the  roorn  temperatures,  which  is  the  same  as  assuming  the  dry-bulb  temperature 
of  the  entering  air  to  be  68  F,  calculate  the  required  air  capacity  of  the  system. 

159 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

Solution. 

199,736  X  55.2 


60  X  12 


minute 


If  a  system  similar  to  the  one  shown  in  Fig.  1  is  used,  1146  Ib  per  minute  will  be  the 
capacity  of  the  dehumidifier  as  well  as  of  the  fan  equipment. 

Example  2.  If  in  addition  to  the  199,736  Btu  per  hour  sensible  heat  load,  the  con- 
ditioned space  has  a  moisture  gain  of  384,000  grains  per  hour,  calculate  the  apparatus 
dew  point  required  to  give  maintained  conditions  of  SO  F  dry-bulb  and  65  F  wet  -bulb, 
with  a  corresponding  56  %  F  dew  point. 

Solution.  With  384,000  grains  of  moisture  per  hour  to  be  picked  up,  the  entering  dew- 
point  temperature  should  be  low  enough  so  that  the  addition  of  this  moisture  will  not 
increase  the  dew  point  above  56  J^  F. 

Grains  per  pound  of  air  saturated  at  56  H  F 
(Table  5,  Chapter  1)  6R.1 

384,000 
Less:    Grains  per  pound  to  be  picked  up,  1146  X  60*  '^ 

Grains  per  pound  allowable  in  entering  air  02.5 

This  corresponds  to  an  apparatus  dew-point  temperature  of  54.17  F. 
Example  8.    Illustration  of  the  by-pass  system.     (See  Fig.  2.J 

Assume  the  same  data  as  for  Example  2.  Instead  of  passing  all  of  the  air  through  the 
dehumidifier  for  cooling  and  dehumidifying,  a  portion  may  be  passed  through  ^  and  the 
balance  be  mixed  with  the  conditioned  air  at  the  leaving  end  of  the  dehumidifier,  the 
mixture  being  proportioned  so  that  the  resultant  conditions  will  be  those  required  to 
give  proper  conditions  in  the  area  considered. 

Solution.  The  quantity  of  air  to  be  dehumidified,  the  quantity  to  be  by-passed,  and 
the  apparatus  dew-point  temperature  may  be  calculated  as  follows: 

Let 

X  —  percentage  of  air  to  be  by-passed. 
Y  =  percentage  of  air  to  be  passed  through  the  dehumidifier. 
/3  —  apparatus  dew-point  temperature,  degrees  Fahrenheit. 

The  quantity  X  of  80-F  air  must  mix  with  the  quantity  Y  of  dehumidified  air  to 
produce  air  with  a  resultant  65  F  wet-bulb  temperature.  Also,  X  quantity  of  air  at 
56  M  F  dew  point  must  be  mixed  with  K  quantity  of  dehumidified  air  to  give  a  resultant 
apparatus  dew-point  temperature  of  54.17  F.  It  is  assumed  that  the  air  passing  through 
the  dehumidifier  is  saturated. 

Solving  simultaneous  equations, 

80.0Z  -f  Ytd  =  68.00  (3) 

56.5JT  +  Ytd  =  54.17  ^ 

23.5J*T  +  0      =  13.83 

x  =  13-803  *  10°  =  59  per  cent,  air  by-passed. 


Y  -  100  —  X        =41  per  cent,  air  passed  through  washer. 

The  second  step  is  to  determine  the  apparatus  dew-point  temperature.    Substitute  X 
in  either  Equation  3  or  Equation  4,  and  solve  for  id  : 

80  X  0.59  +  /d  X  0,41  =  68 

gg    _    AJ 

/d  =         •  -  —  =  51.2  F,  the  apparatus  dew  point. 

0.41 

160 


CHAPTER  9 — CENTRAL  Am  CONDITIONING  SYSTEMS 


HEAT  TO  BE  REMOVED  BY  COOLING  AND  DEHUMIDIFYING 

APPARATUS 

Example  4-  Assume  the  same  data  as  for  Example  3.  If  the  amount  of  outside  air,  at 
95  F  dry-bulb  and  75  F  wet-bulb,  required  for  ventilation  has  been  found  to  be  169  Ib 
per  minute,  determine  the  refrigeration  capacity  required. 

Solution.  As  the  total  weight  of  the  air  introduced  per  minute  is  1146  Ib,  and  41  per 
cent  of  it  goes  through  the  dehumidifier,  the  total  work  to  be  done  may  be  computed 
as  follows: 

Air  passing  through  humidifier,  1146  X  0.41 470  Ib 

Less:    Outside  air  for  ventilation 169  Ib 


Return  air 301  Ib 

The  refrigeration  required  for  the  return  air  is: 

Total  heat  per  pound  at  65  F 29.65  Btu 

Less:    Total  heat  per  pound  at  51.2  F 20.85  Btu 


Requirement  for  cooling  1  Ib  of  return  air 8.80  Btu 

301  Ib  X  8.80  Btu   =  2649  Btu  per  minute  required  to  coo!  the 
return  air. 

The  refrigeration  required  for  the  outside  air  is: 

Total  heat  per  pound  of  outside  air 37.81  Btu 

Less:    Total  heat  per  pound  at  51.2  F 20.85  Btu 


Requirement  to  cool  1  Ib  of  outside  air 16.96  Btu 

169  Ib  X  16.96  Btu  =  2866  Btu  per  minute  required  to  cool  the 
outside  air. 

Thus,  the  total  refrigeration  required  is: 

2649  Btu  -f-  2866  Btu  =  5515  Btu  per  minute,  which  is  equivalent 
to  a  load  of  27.6  tons  of  refrigeration. 

SIZE  OF  REHEATERS 

A  properly  designed  air-conditioning  system  will  have  reheaters  of 
sufficient  capacity  to  heat  the  conditioned  air  from  the  apparatus  dew- 
point  temperature  to  the  outlet  delivery  temperature.  If  winter  heating 
is  to  be  accomplished,  consult  Chapter  22. 

The  following  general  formula  may  be  used  to  determine  the  amount  of 
heat  necessary  to  reheat  a  given  quantity  of  air: 

H\  =  0.24  (ty  -  /d)  M  (5) 

where 

H\  =  heat  to  be  supplied  to  reheater  coil,  Btu  per  hour. 

Example  5.  Assume  the  same  data  as  for  Example  1,  and  find  the  amount  of  reheating 
required. 

Solution. 

H\  =  0.24  (68  -  54.17)  1146  X  60  =  228,200  Btu  per  hour. 

SURFACE  COOLING  PROBLEM 

The  amount  of  coil  surface  required  for  a  given  amount  of  work  is 
dependent  upon  factors  previously  listed.  Obviously,  the  various  types  of 
surfaces  made  available  by  different  manufacturers  will  have  different 

161 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

transmission  values.  It  is  recommended  that  the  designer  consult  the 
latest  manufacturers'  catalogs  because  more  accurate  ratings  are  being 
issued  from  time  to  time. 


Airo 
60  Fdr 

ut 
/-bulb 

C 

3 

C 

Air  m 
95  F  dry-bulb 
78  F  wet-  bulb 

Water  in 

~~    50  F 


_^  Water  out 
'50  F  +  30F  =  80  F 


FIG.  4.    COUNTER-FLOW  SURFACE  COOLING  DIAGRAM 

Example  6.  It  is  desired  to  cool  and  dehumidify  30,000  cfm  of  air  at  95  F  dry-bulb, 
78  F  wet-bulb,  and  72  F  dew  point,  to  a  60  F  dew  point.  Cooling  water  is  available  at 
50  F  in  a  quantity  which  will  allow  a  30  F  rise  in  temperature  to  be  used.  The  counter- 
flow  surface  cooling  used  is  sketched  in  Fig.  4. 

Solution.  The  pounds  of  partially  saturated  air  cooled  and  dehumidified  per  hour 
equal  60  times  the  cubic  feet  of  air  at  95  F  dry-bulb  and  78  F  wet-bulb  brought  past  the 
coil  surface  per  minute,  multiplied  by  the  pounds  per  cubic  foot  of  the  air  as  determined 
from  Table  3,  Chapter  1. 

30,000  X  60  X  0.0708  =  127,440  Ib  per  hour.  . 

The  total  heat  Ht  to  be  removed  per  hour  by  the  surface  coil  is  found  to  be  equal  to 
the  pounds  of  partially  saturated  air  passed  over  the  coil  per  hour  times  the  difference 
between  the  total  heat  of  air  at  78  F  wet-bulb  and  at  60  F  wet-bulb. 

Ht  =  127,440  (40.64  -  26.18)  =  1,842,000  Btu  per  hour. 

The  latent  heat  H\  to  be  removed  per  hour  will  be  found  by  multiplying  the  pounds  of 
partially  saturated  air  passed  over  the  coils  per  hour  by  the  difference  in  the  latent  heat 
of  the  air  per  pound  at  the  initial  and  final  dew  points. 

Hi  =  127,440  (17.79  -  11.69)  =  777,000  Btu  per  hour. 

The,  sensible  heat  Hg  to  be  removed  per  hour  is  equal  to  the  total  heat  of  the  air  less 
its  latent  heat. 

Hs  =  Ht  -  Hi  «  1,842,000  -  777,000  =  1,065,000  Btu  per  hour. 

Manufacturers'  standard  ratings  for  surface  coolers  are  usually  based 
on  the  cubic  feet  of  air  passed  through  their  equipment  per  minute, 
reduced  to  the  conditions  of  saturated  air  measured  at  a  temperature  of 
70  F.  In  the  present  example,  to  convert  the  127,440  Ib  of  air  cooled  per 
hour  to  a  basis  which  will  permit  the  use  of  such  standard  ratings,  it  is 
necessary  to  multiply  the  pounds  of  air  cooled  per  hour  by  the  specific 
volume  of  the  air,  and  to  divide  by  60. 

127>44°  *  13'69  -  29,100  cfm  of  70  F  saturated  air. 
ou 

The  amount  of  cooling  water  necessary  when  a  30  degree  rise  in  its 
temperature  is  to  be  used  is: 

'.  1,842,000 


30.  X  8.34  X  60 
162 


=  123  gpm. 


CHAPTER  9 — CENTRAL  AIR  CONDITIONING  SYSTEMS 


With  counter  flow  of  air  and  water,  it  is  necessary  to  determine  the 
mean  temperature  difference  between  the  air  and  the  water  in  order  to 
properly  use  the  transmission  coefficients  given  in  apparatus  rating  tables. 

J}^   _  £>2 

Mean  temperature  difference  =  -  ^—  (6) 

Ioge  B; 

where 

Dl  =  the  difference  between  the  temperatures  of  inlet  air  and  outlet  water,  degrees 

Fahrenheit. 

Do  =  the  difference  between  the  temperatures  of  outlet  air  and  inlet  water,  degrees 
Fahrenheit. 

(95  -  80)  -  (60  -  50)  _ 

—  ;  —  (95  -  so)  --  12-33  R 

loge  (60  -  50) 

If  from  apparatus  rating  tables  based  on  air  velocities  over  the  coils  and 
water  velocities  through  the  coils,  it  has  been  found  that  the  transmission 
coefficient  is  equal  to  8.0  Btu  per  square  foot  per  degree  difference  in 
mean  temperature  between  the  air  and  the  water,  the  area  of  cooling  coil 
surface  necessary  will  be  equal  to  the  sensible  heat  divided  by  the  trans- 
mission coefficient  and  also  by  the  mean  temperature  difference. 


•«   f\fiK  AA/"l 

'  ^  VWoo    =  10,800  square  feet  of  cooling  coil  surface  necessary. 
o.U 


The  latent  heat  is  taken  out  at  the  same  time  the  sensible  heat  is 
extracted,  but  no  extra  surface  is  required  unless  the  latent  heat  exceeds 
approximately  40  per  cent  of  the  total  heat.  This  is  because  the  wetted 
surface  has  a  much  higher  coefficient  of  transmission.  Approximately 
10  per  cent  more  surface  should  be  added  if  the  latent  heat  exceeds  40  per 
cent  of  the  total  heat. 

AUXILIARY  EQUIPMENT 

Consult  Chapters  14,  17,  19,  20,  and  22  for  information  on  the  air 
distribution  system;  air  outlets  and  inlets;  static  pressure  on  fan;  fan 
motor,  and  drive;  and  the  control  system. 

PROBLEMS  IN  PRACTICE 

1  •  In  summer  air  conditioning  what  factors  control  the  difference  between 
the  dry-bulb  temperature  of  the  conditioned  space  and  the  dry-bulb  tem- 
perature of  the  entering  air? 

1.  The  duct  and  supply  grille  arrangement  permitted  by  architectural  and  structural 
requirements  for  the  particular  space,  e.g.,  ceiling  height  and  obstructions  on  ceilings, 
such  as  beams. 

2.  The  state  of  activity  of  the  occupants. 

3.  The  outlet  velocity  at  the  grille,  as  limited  by  noise  level  requirements. 

4.  The  direction  of  the  jet  relative  to  the  occupants. 

5.  In  some  cases,  the  temperature  of  the  available  water  supply,  which  may  have  some 
bearing  on  the  air  delivery  temperature. 

2  •  What  factors  determine  the  volume  of  conditioned  air  which  must  be 
delivered  to  the  space? 

The  sensible  heat  to  be  removed,  and  the  allowable  temperature  differential. 

163 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

3  •  What  factors  determine  the  dew  point  of  the  air  entering  the  space? 

The  maximum  dew  point  desired  in  the  conditioned  space,  and  the  moisture  gain  in  the 
space  per  unit  weight  of  air  supplied. 

4  •  Why  must  the  air  leaving  a  dehumidifying  type  air  washer  be  reheated 
before  delivery? 

The  air  leaves  the  dehumidifying  air  washer  saturated  at  a  relatively  low  temperature 
which  in  most  cases  is  lower  than  the  allowable  delivery _dry-bulb  temperature.  Also, 
the  air  may  possibly  be  carrying  a  small  amount  of  entrained  water  which  might  settle 
out  in  the  ducts  near  the  washer  and  cause  corrosion  difficulties. 

5  •  What  methods  are  used  for  reheating  air? 

1.  Passing  it  over  reheating  coils. 

2.  Mixing  it  with  by-passed  air  at  a  higher  temperature. 

6  •  What  determines  the  final  temperature  of  the  spray  water  in  a  dehumidifier? 

Because  of  the  effectiveness  of  the  heat  transfer  between  air  and  finely  divided  spray 
water  in  a  well  designed  dehumidifier,  the  air  will  be  cooled  to  within  1  or  2  F  of  the  final 
water  temperature,  provided  the  air  velocity  through  the  washer  does  not  exceed  600  fpm. 
This  final  temperature  should  then  be  taken  as  1  or  2  F  lower  than  the  required  dew 
point  of  the  air  leaving  the  washer. 

7  •  WThat  are  the  advantages  of  using  counter  flow  of  ah*  and  water  in  surface 
coolers? 

Counter  flow  results  in  a  higher  mean  temperature  difference  than  does  parallel  flow  for 
the  same  range  of  air  and  water  temperatures,  which  means  that  less  cooling  surface  is 
required.  Counter  flow  permits  higher  initial  water  temperatures  and  also  allows  a 
greater  temperature  rise  for  the  water.  These  factors  combine  to  reduce  the  cost  of 
circulating  and  refrigerating  the  cooling  water. 

8  •  What  factors  other  than  cost  should  be  considered  in  determining  whether 
to  use  a  central  system  or  another  type? 

a.  Appearance:    The  equipment  must  be  designed  to  harmonize  with  the  architecture 
of  the  building. 

b.  Distribution:    The  system  must  maintain  adequate  and  uniform  air  motion  over  the 
entire  conditioned  space. 

c.  Control:   The  control  system  must  be  designed  to  give  effective  partial  load  operation. 

9  •  Can  the  central  cooling  and  dehumidifying  system  be  used  as  an  all-year- 
round  conditioner? 

By  modifying  the  control  system  and  adding  blast  coils  or  a  water  heater  to  the  spray 
type  system,  the  cooling  system  will  function  as  one  for  heating  and  humidifying.  The 
surface  cooling  type  may  be  transformed  by  modifying  the  control,  and  adding  another 
set  of  coils  and  a  humidifier. 

10  •  Will  the  tons  of  refrigeration-effect  per  day  be  the  value  calculated  in 
Example  4  of  this  chapter  times  the  hours  of  operation? 

No.  The  tons  of  refrigeration-effect  are  functions  of  the  load.  The  components  of  the 
load  vary,  that  is,  the  number  of  people  occupying  the  space,  the  outdoor  conditions,  and 
the  solar  radiation  will  change  from  hour  to  hour  and  from  day  to  day.  The  calculated 
load  represents  the  maximum  required  for  design  peak  conditions. 

11  •  Will  the  quantity  of  return  air  required  in  Example  4  of  this  chapter  be 
used  all  season? 

No.  When  the  outdoor  wet-bulb  temperature  becomes  lower  than  the  maintained  wet- 
bulb  temperature,  it  is  more  economical  to  use  all  outside  air  than  to  dehumidify  the 
return  air. 

164 


B 


Chapter  10 

COOLING  METHODS 

Methods  of  Cooling  Air,  Evaporative  Cooling.,  Dehumidification., 

Silica  Gel  System,  Alumina  System,  Design  of  System,  Operating 

Methods,  Steam  Jet  System,  Compressors,  Refrigerants,  Methods 

of  Cooling,  Condensers 

Y  using  any  of  the  following  four  methods,  or  any  combination  of 
them,  effective  temperature  (see  Chapter  2)  may  be  reduced. 

a.  Sensible  cooling:    Lowering  of  the  dry-bulb  temperature  by  the  removal  of  sensible 
heat  without  change  of  the  dew-point  temperature. 

b.  Dehumidifying:    Lowering  of  the  dew-point  temperature  by  the  removal  of  mois- 
ture without  change  of  the  dry-bulb  temperature. 

c.  Evaporative  cooling:    Lowering  of  the  dry-bulb  temperature  through  the  evapor- 
ation of  moisture  without  the  addition  or  the  subtraction  of  heat. 

d.  Air  motion:     Increasing  the  air  motion  over  the  body  with  the  resulting  higher 
evaporation  from  the  skin. 

As  an  example,  let  the  condition  be  considered  of  92  F  dry-bulb,  with  a 
40  per  cent  relative  humidity,  corresponding  to  a  wet-bulb  temperature  of 
72.8  F,  and  an  effective  temperature  for  still  air  of  81.1  F.  This  effective 
temperature  may  be  reduced  3.1  F  by  any  of  the  four  basic  methods 
mentioned,  as  follows : 

First,  by  lowering  the  dry-bulb  temperature  to  85.5  F  without  changing  the  dew-point 
of  64.2 ;  this  gives  an  effective  temperature  of  78  F. 

Second,  by  reducing  the  moisture  content  of  the  air  to  46  grains  per  pound  of  dry  air 
without  changing  the  dry-bulb  temperature;  this  gives  an  effective  temperature  of  78  F. 

Third,  by  reducing  the  dry-bulb  temperature  to  83.8  F  without  changing  the  total 
heat  of  the  air.  This  requires  the  evaporation  of  14  grains  of  moisture  per  pound  of  dry 
air,  and  the  effective  temperature  will  become  78  F. 

Fourth,  by  increasing  the  air  movement  from  still  air  to  460  fpm,  a  velocity  which  will 
reduce  the  effective  temperature  3.1  F  from  81.1  F  to  78  F. 

Method  to  Employ 

The  best  method  of  reducing  the  effective  temperature  in  any  specific 
case  will  depend  on  the  accompanying  circumstances  and  can  be  deter- 
mined only  by  a  thorough  analysis  made  by  a  competent  engineer. 
Generally  speaking,  the  removal  from  the  air  of  the  sensible  heat,  or 
moisture,  or  both,  by  sensible  cooling  or  dehumidifying  is  the  most 
satisfactory  method.  Adequate  results  by  the  utilization  of  air  motion  or 
by  evaporative  cooling  are  difficult  to  obtain  because  of  the  dependence 
of  both  methods  upon  climatic  conditions  beyond  the  engineers'  control 
although  these  methods  are  much  less  expensive  than  the  first  two 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

mentioned.  Cooling  by  evaporation  is  satisfactory  only  when  the  air  to 
be  cooled  is  very  dry ;  air  motion  as  a  means  of  producing  cooling  effect  is 
never  entirely  adequate  in  the  range  of  high  temperatures.  Of  the  two, 
evaporative  cooling,  or  adiabatic  saturation  of  the  air,  is  a  much  more 
dependable  method  which  will  make  more  reduction  in  the  effective 
temperature  than  will  an  increasing  air  motion  within  permissible  limits. 
As  an  example  of  this,  consider  an  outdoor  condition  of  96  F  dry-bulb 
and  80  F  wet-bulb.  The  effective  temperature  is  85.7  F  and,  if  the  still 
air  is  moved  with  a  velocity  of  300  fpm,  the  effective  temperature  will  be 
reduced  only  2.0  F  while  saturation  at  the  wet-bulb  temperature  would 
reduce  the  effective  temperature  5.7  F.  At  300  fpm  velocity  this  satu- 
rated air  would  reduce  the  effective  temperature  to  75.6  F,  thus  making  a 
total  improvement  of  10.1  F. 

Evaporative  Cooling 

Evaporative  cooling  is  accomplished  by  passing  air  through  a  water 
spray  in  which  the  water  is  being  continually  recirculated.  The  air, 
entering  in  an  unsaturated  condition,  evaporates  a  part  of  the  water  at  the 
expense  of  the  sensible  heat  As  this  is  an  adiabatic  transfer,  the  total 
heat  content  of  the  air  remains  constant,  while  the  dew  point  rises  and  the 
dry-bulb  falls  until  the  air  is  saturated.  A  system1  of  ducts  and  a  propel- 
ling fan  are  used  to  distribute  the  air  in  a  proper  manner. 

It  will  be  seen  that  the- reduction  in  dry-bulb  temperature  is  a  direct 
function  of  the  wet-bulb  depression  of  the  air  entering  the  ddhumidifier 
and  that  the  resulting  air  temperature  is  governed  entirely  by  the  entering 
wet-bulb  temperature  of  the  outside  air. 

Dehumidification 

Dehumidification  may  be  accomplished  in  three  ways: 

1.  By  cooling  the  air  below  the  dew  point  and  causing  a  part  of  the  moisture  contained 
to  precipitate. 

2.  By  extracting  all  or  part  of  the  moisture  by  absorption. 

3.  By  extracting  all  or  part  of  the  moisture  by  adsorption. 

As  used  in  this  discussion,  the  term  adsorption  pertains  to  the  action  of 
a  substance  in  condensing  a  gas  or  vapor  and  holding  the  condensate  on 
its  surface  without  any  change  in  the  chemical  or  physical  structure  of  the 
substance  and  with  the  release  of  sensible  heat.  The  term,  absorption, 
implies  a  change  in  the  chemical  or  physical  structure  of  a  substance  in  the 
process  of  dehydrating  air.  Adsorbers  include  silica  gel  and  lamisilite; 
absorbers  include  sulphuric  acid. 

Dehumidification  by  Refrigeration 

Air  conditioning  imposes  requirements  on  refrigeration  equipment  not 
usually  found  in  general  cooling  work,  so  that  specially  designed  apparatus 
is  often  needed  to  replace  that  normally  used  for  industrial  cooling. 
Standard  equipment  can  be  adapted  to  meet  air  conditioning^  require- 
ments but  extreme  care  must  be  taken  to  determine  the  limits  of  its 
applicability. 

!See  Air  Washer  Performance  in  Chapter  11;  also  Theory  of  Atmospheric  Cooling  in  same  chapter. 

166 


CHAPTER  10 — COOLING  METHODS 


In  Industrial  or  process  cooling  systems  the  load  is  fairly  constant,  noise 
in  operation  is  not  of  paramount  importance,  space  is  available  or  ^re- 
latively cheap,  condenser  water  is  not  a  source  of  worry,  and  the  cooling 
system  is  to  a  great  extent  separate  and  independent  of  other  mechanical 
equipment.  By  contrast,  air  conditioning,  especially  as  used  for  space 
cooling  and  comfort  work  in  office  buildings,  theaters,  and  places  where 
people  gather  requires  special  consideration  of  all  these  factors.  Space  in 
public  buildings  is  limited  and  condenser  water  is  expensive.  Noise 
interferes  with  the  occupants,  and  the  cooling  equipment  must  dovetail 
with  the  other  air-handling  apparatus.  Most  important,  the  load  fluctu- 
ates tremendously  and  is  seasonal. 


Heat  of  Compression 
Added  to  Gas 


Low  Pressure  Saturated 

X 

Hot  In              . 

Gas 

Compressor 

High-F 

Vessure 
Condei 

Superheated  Gas 
*"•       Cold  in 

Evaporator  or  Cooler 

Heat  Added  to 
Refrigerant  by 
Substance  Cooled 

1= 

Refrigerant  by 

Cold  Out                       £,              .     ..  . 
£\ExpansK)n  Valve 

for  Reducing  Pressure                                          . 

Hot  Out 

High  Pressure  Saturated  Liquid 
FIG.  1.    TYPICAL  REFRIGERATION  DIAGRAM 


A  complete  discussion  of  the  thermodynamic  problems  of  refrigeration 
is  given  In  the  Refrigerating  Data  Book2, 1934,  so  only  a  brief  description 
of  the  cycle  will  be  given  here  before  the  problems  peculiar  to  air  con- 
ditioning are  considered. 

The  refrigeration  system  consists  of  three  main  parts,  the  evaporator, 
the  condenser,  and  the  compressor.  Fig.  1  shows  a  diagram  of  the  cycle. 
Heat  is  absorbed  in  the  evaporator  and  released  in  the  condenser.  The 
compressor  changes  the  level  of  the  heat  by  taking  it  from  a  lower  to  a 
higher  plane.  There  are  also  many  valves,  accessories,  and  special  devices 
necessary  for  proper  operation,  which  vary  somewhat  with  different  types 
of  cooling  systems  and  different  refrigerants. 

In.  a  simple  illustrative  cycle  of  a  refrigeration  system,  the  liquid 
refrigerant  under  high  pressure  has  both  its  pressure  and  temperature 
reduced  by  being  expanded  through  a  suitable  valve  into  an  evaporator  or 
cooler.  Within  the  evaporator  the  low  temperature  of  the  refrigerant 
allows  it  to  absorb  heat  from  the  substance  to  be  cooled,  which  surrounds 
the  eyaporator.  This  absorption  of  heat  increases  the  pressure  of  the 


*PttbSsfced  by  American  Society  of  Refrigerating  Engineers. 

167 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


refrigerant,  and  a  compressor  is  employed  to  withdraw  enough  low- 
pressure  saturated  gas  to  keep  the  cooling  action  of  the  evaporator  con- 
tinuous. The  withdrawn  gas  is  discharged  from  the  compressor  to  the 
condenser  in  the  form  of  a  high-pressure  superheated  gas  which  includes 
the  heat  added  through  its  compression.  In  the  condenser,  because  heat 
is  taken  from  the  gaseous  refrigerant  by  the  condensing  medium,  usually 
water,  the  refrigerant  again  becomes  the  high-pressure  saturated  liquid 
with  which  the  cycle  started. 

The  cooling  water,  which  may  come  from  a  deep  well  or  from  a  city 
main,  may  be  utilized  for  some  purpose  after  it  has  been  warmed  a  few 
degrees  in  the  condenser,  or  after  use  it  may  be  exposed  to  the  atmosphere 


J3RY  AIR 


< 


COOLER 


ADSORPTION 


FAN 


ACTIVATION 


FAN 


WET 


GAS   HEATER 


FIG.  2.    SILICA  GEL  AIR-CONDITIONING  SYSTEM — SINGLE  STAGE  ADSORPTION 

in  a  spray  pond  or  cooling  tower  and  have  its  temperature  reduced  to  a 
point  where  the  water  may  be  used  again.    (See  Chapter  11.) 

Silica  Gel  System 

Silica  gel  is  a  chemical  composition  made  from  sodium  silicate  and  acid, 
the  chemical  formula  being  SiO2.  It  has  an  appearance  greatly  resembling 
that  of  clear  quartz  sand  but  it  differs  in  structure  in  that  the  crystals 
are  highly  porous,  with  voids  constituting  41  per  cent  by  volume  although 
the  pores  are  microscopic  in  size.  This  material  possesses  the  property  of 
being  able  to  adsorb  a  substantial  portion  (about  25  per  cent  of  its  own 
weight)  of  moisture  from  the  air  without  any  increase  in  its  volume. 
After  the  silica  gel  has  become  "  saturated  "  or  has  adsorbed  moisture  to 
the  limit  of  its  capacity,  the  moisture  may  be  driven  from  it  by  the 
application  of  heat,  again  without  change  in  the  structure,  volume,  or 
chemical  composition  of  the  silica  gel.  This  cycle  may  be  repeated  in- 
definitely. When  applied  to  air  conditioning  the  silica  gel  which  is 
exposed  to  the  air  reduces  the  moisture  content  in  the  air  and  releases 
sensible  heat  which  may  be  readily  removed  from  the  air.  A  typical 
diagram  is  shown  in  Fig.  2. 

168 


CHAPTER  10 — COOLING  METHODS 


Practical  Application  of  Silica  Gel 

Silica  gel  has  two  applications  when  used  to  replace  refrigeration.  In 
the  one  principally  used,  the  air  from  which  moisture  is  to  be  extracted  is 
taken  through  silica  gel  beds  by  suction  or  pressure  fans,  and  by  means  of 
this  process  the  moisture  becomes  adsorbed  by  the  silica  gel  and  the  air 
leaves  at  a  lower  dew  point  and  a  higher  sensible  temperature  than  those 
at  which  it  entered.  If  this  air  is  passed  over  surface  coolers  in  which  tap 
water  or  another  cooling  medium  is  flowing  through  tubes,  a  certain 
amount  of  sensible  heat  will  be  removed.  The  air  leaves  the  surface  cooler 
or  interchanger  with  the  same  dew  point  with  which  it  emerged  from  the 
silica  gel  beds,  but  with  a  lower  dry-bulb  temperature,  although  the  dry- 
bulb  temperature  may  be  higher  than  the  temperature  of  the  air  entering 
the  silica  gel  beds. 

In  another  method,  the  first  two  of  the  steps  outlined  are  duplicated, 
and  in  addition  the  air  is  carried  through  a  spray  type  washer.  Because 
the  air  enters  the  washer  with  a  low  wet-bulb,  and  because  adiabatic 
saturation  will  take  place  at  a  temperature  close  to  the  entering  wet-bulb, 
considerable  cooling  of  the  air  can  be  accomplished;  but  this  can  be  done 
only  with  a  consequent  increase  of  the  dew  point. 

It  is  necessary  to  reactivate  the  silica  gel  after  it  has  adsorbed  about 
25  per  cent  of  its  own  weight  in  the  form  of  moisture.  As  reactivation 
requires  a  high  temperature  and  since  silica  gel  is  only  active  at  low  tem- 
peratures, cooling  of  the  beds  must  also  be  completed  before  they  can  be 
used  again.  This  necessitates  three  stages  in  the  silica  gel  containers  and 
requires  either  three  beds  of  silica  gel  or  one  bed  divided  and  automatically 
put  in  position.  The  reactivation  is  usually  done  by  means  of  gas  or  oil 
fires  and  the  cooling  of  the  beds  by  means  of  indirect  water  cooling  or  by 
means  of  small  quantities  of  dehydrated  air  taken  from  the  system  beyond 
the  interchanger. 

Alumina  System  of  Adsorption 

Activated  alumina  contains  a  trifle  over  91  per  cent  of  aluminum 'oxide, 
AlzOz,  which  material  will  adsorb  nearly  100  per  cent  of  the  vapor  in  the 
air  up  to  about  8  or  10  per  cent  of  the  weight  of  the  adsorbing  material, 
after  which  the  adsorption  falls  off  gradually  as  the  saturation  point  is 
approached.  The  application  is  quite  similar  to  that  employed  for  silica 
gel;  that  is,  the  material  is  exposed  to  the  air  flow  and  after  reaching 
about  75  per  cent  saturation  is  reactivated  by  removing  the  moisture 
adsorbed  by  means  of  applied  heat.  The  actual  scheme  generally  fol- 
lowed in  the  use  of  this  material  for  continuous  service  varies  somewhat 
from  silica  gel  inasmuch  as  the  material  is  placed  in  three  units  which  are 
used  consecutively  for  the  different  steps.  These  steps  permit  each  unit 
to  operate  as  follows : 

a.  In  series  with  the  preceding  unit. 

b.  Alone. 

c.  In  series  with  the  following  unit. 

This  plan  allows  for  adsorption,  reactivation,  and  cooling,  in  a  manner 
similar  to  that  used  with  silica  gel. 

Taking  a  single  unit,  when  it  is  in  the  a  step  and  operating  with  the 

169 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

preceding  unit,  the  alumina  adsorbs  approximately  25  per  cent  of  the 
moisture  in  the  air  and  takes  up  about  1.3  per  cent  of  its  weight  of  water. 
During  the  second  step  when  it  is  operating  alone,  it  takes  up  100  per  cent 
of  the  moisture  in  the  air  until  the  weight  of  the  water  adsorbed  is  brought 
up  to  about  6.7  per  cent.  During  the  third  step  when  the  unit  is  operating 
with  the  succeeding  unit,  it  extracts  about  75  per  cent  of  the  moisture  in 
the  air  until  the  water  weight  adsorbed  comes  up  to  about  10  per  cent  of 
the  weight  of  the  adsorber.  The  time  allowable  for  reactivating  is  equal 
to  the  time  occupied  by  the  second  unit  adsorbing  alone,  plus  the  time 
when  the  second  and  third  units  are  adsorbing  in  series,  plus  the  time 
when  the  third  unit  is  adsorbing  alone,  at  the  expiration  of  which  time  the 
first  unit  will  be  again  required. 

The  temperature  of  air  used  for  alumina  reactivation  is  usually  between 
300  and  700  F  and  the  air  flow  rate  will  have  to  be  higher  with  the  low 
temperature  air  than  it  will  be  with  reactivating  air  of  higher  temperature. 
For  example,  air  at  400  F  for  reactivating  will,  at  10  cu  ft  per  hour  per 
pound  of  alumina,  require  about  6  hours  for  reactivation.  In  the  three 
unit  system,  after  reactivation  the  cooling  of  the  activated  alumina  may 
be  carried  out  with  considerable  rapidity  by  using  dry  air  from  the  adsorp- 
tion unit  for  circulation  through  the  unit  which  has  just  completed  reacti- 
vation. The  final  temperature  of  the  unit  before  it  goes  back  into  service 
should  be  not  over  200  F.  As  a  basis  for  computing  the  amount  of  cooling 
air  required  for  reactivation,  each  cubic  foot  of  cooling  air  has  been  found 
capable  of  removing  2.2  Btu  when  heated  from  85  to  200  F  and  of  provid- 
ing a  sufficient  margin  of  safety  in  operation. 

Design  of  System 

When  designing  air  conditioning  systems,  the  capacity  of  equipment  is 
decided  by  selecting  apparatus  of  sufficient  size  to  maintain  predetermined 
temperatures  and  humidities  in  treated  spaces  when  arbitrarily  estab- 
lished maximum  atmospheric  temperatures  occur  coincident  with  given 
conditions  of  population,  lighting,  and  power  consumption.  These  factors 
determine  the  maximum  duty  of  the  cooling  system.  The  duty  does  not 
necessarily  determine  the  size  or  capacity  of  the  refrigeration  apparatus. 
The  refrigerating  capacity  is  expressed  in  tons,  each  ton  being  equal  to  the 
absorption  of  the  heat  given  up  by  one  ton  of  ice  at  32  F  melting  to  water 
at  32  F  in  24  hours.  This  is  equivalent  to  heat  absorption  at  a  rate  of 
approximately  200  Btu  per  minute,  or  12,000  Btu  per  hour. 

After  the  maximum  duty  is  determined,  the  other  factors  concerning 
the  installation  must  be  investigated.  The  total  heat  to  be  removed  by 
the  cooling  system  has  many  sources,  some  substantially  constant  and 
others  extremely  variable.  These  sources  can  be  roughly  classified  as 
follows,  the  first  column  indicating  the  order  in  amount  and  the  second 
the  order  in  variability: 

1.  Fresh  air  supplied.  1.  Fresh  air  supplied. 

2.  Population.  2.  Transmission  through  the  structure. 

3.  Transmission  through  the  structure.  3.  Light  and  power  consumed. 

4.  Light  and  power  consumed.  4.  Population. 

By  combining  these  two  columns,  a  third  grouping  is  obtained  -  as 
follows: 

170 


CHAPTER  10 — COOLING  METHODS 


1.  Fresh  air  supplied.  3.  Population. 

2.  Transmission  through  the  structure.  4.  Light  and  power  consumed. 

In  this  last  arrangement,  the  first  two  items  are  governed  by  atmos- 
pheric conditions  and  they  are  therefore  subject  to  tremendous  fluctu- 
ations in  value.  As  they  generally  form  40  to  60  per  cent  of  the  entire 
maximum  load,  the  duty  of  the  cooling  system  will  be  much  less  than 
maximum  most  of  the  time. 

The  transmission  through  the  structure  is  especially  influenced  by  the 
sun.  (See  Chapter  8.)  In  many  cases,  because  of  the  heat  flow  resistance 
of  the  structure,  the  heat  from  the  sun  is  retarded  until  it  is  compensated 
for  by  a  reduced  general  temperature  out-of-doors. 

A  survey  of  Weather  Bureau  records  indicates  that  maximum  tempera- 
tures occur  less  than  5  per  cent  of  the  cooling  period  and  also  that  the 
duration  of  peak  conditions  is  never  more  than  three  or  four  hours. 

Two  factors  control  the  size  of  the  refrigeration  system,  the  evaporator 
or  suction  temperature,  and  the  condenser  or  head  temperature.  With 
the  knowledge  that  the  system  will  operate  most  of  the  time  with  a  load  of 
not  over  60  per  cent  of  maximum,  and  that  maximum  demands  will  occur 
infrequently  and  only  for  short  periods,  some  provision  must  be  made  to 
insure  economical  operation  under  average  conditions.  This  can  be  done 
by  overloading  the  machine  under  extreme  demands  and  basing  the  design 
on  normal  or  average  loads.  Flexibility  in  arrangement  can  be  provided 
in  several  ways. 

Variations  in  load  change  the  efficiency  of  any  machine  and  a  refrigera- 
ting system  can  be  costly  and  inefficient  if  improperly  designed  or  operated. 
Fortunately,  the  trouble  can  be  concentrated  in  the  compressor  and  the 
problem  relieved  of  many  complications.  It  is  comparatively  easy  to 
furnish  condensers  and  evaporators  to  carry  the  maximum  load  so 
arranged  that  they  will  function  properly  at  small  demands.  They  affect 
the  compressor  performance  to  some  extent  but  most  of  the  compressor 
problems  are  in  the  machine  itself. 

Variations  in  load  are  usually  effected  by  lowering  the  suction  tem- 
perature and  pumping  a  larger  volume  of  gas  per  ton  through  a  greater 
pressure  range.  This  is  possible  because  the  latent  heat  of  the  refrigerant 
remains  nearly  constant  throughout  the  small  range  used  and  the  specific 
volume  varies  rapidly  with  change  in  pressure.  As  the  compressor  must 
remove  the  refrigerant  evaporated,  the  evaporator  temperature  fixes  the 
displacement  required.  The  objection  to  such  method  is  that  the  total 
power  consumed  remains  nearly  constant  and  the  power  per  unit  of 
cooling  increases  rapidly  as  the  total  output  is  reduced.  Such  operation 
is  satisfactory  as  long  as  the  load  is  kept  within  10  per  cent  of  the  rating 
of  the  compressor  but  this  condition  does  not  commonly  occur  in  air 
conditioning  applications. 

Operating  Methods 

It  is  possible  to  divide  the  entire  refrigeration  system  into  a  number  of 
small  units,  which  will  allow  cutting  in  and  out  of  compressors  and  con- 
densers as  the  load  fluctuates.  This,  however,  is  an  expensive  method  as 
a  number  of  small  units  are  usually  more  expensive  than  one  large  unit. 
There  is  a  certain  amount  of  duplication  of  equipment  necessary,  which 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

tends  to  increase  the  initial  cost  of  the  system  and  which  makes  the  fixed 
charges,  applicable  to  the  operation  of  the  air  conditioning  and  cooling 
system,  greater  than  necessary. 

A  second  method  of  providing  for  economy  of  operation  is  to  have 
storage  capacity  which  can  be  utilized  during  the  peak  period.  A  further 
reference  to  the  Weather  Bureau  records  indicates  that  maximum  con- 
ditions prevail  during  the  day  for  not  more  than  three  hours,  and  con- 
sequently the  refrigerating  system  can  be  run  for  a  longer  period  at 
maximum  efficiency  with  tanks  to  store  cold  water  or  brine  for  supple- 
menting the  actual  output  of  the  refrigerating  equipment  when  the  load  is 
more  than  the  machine  will  carry.  This  situation  brings  complications. 
Storage  tanks  require  space  and  extra  apparatus,  which  increase  the  cost 
of  the  entire  system,  and  further,  it  is  difficult  to  determine  what  the  size 
of  the  compressor  should  be  because  of  the  other  variables  which  enter  the 
problem.  Depending  upon  the  availability  of  storage  space,  the  com- 
pressor could  be  designed  for  any  reasonable  percentage  of  the  maximum 
load,  so  the  smaller  the  compressor,  the  larger  the  storage  space,  and 
vice  versa. 

A  third  method  is  to  provide  in  the  compressor  itself  some  means  of 
reducing  the  capacity.  This  can  be  done  by  varying  the  speed  and  con- 
sequently the  displacement  of  the  compressor,  or  by  varying  the  dis- 
placement, either  by  a  partial  by-pass  of  the  cylinder  or  by  a  clearance 
pocket  in  the  head  of  the  cylinder  when  reciprocating  compressors  are 
used.  It  might  be  assumed  that  the  efficiency  would  remain  practically 
constant.  This  is  not  correct,  inasmuch  as  the  machine  friction  remains 
constant  with  the  by-pass  or  clearance  pocket  method  and  this  raises  the 
power  required  per  ton  of  refrigeration  developed.  Also,  the  volumetric 
efficiency  of  the  machine  falls  off  rather  rapidly  when  the  clearance  pocket 
or  partial  by-pass  is  used.  By  varying  the  speed  of  the  compressor,  the 
efficiency  of  the  power  unit  falls  off  as  the  speed  is  reduced,  while  the 
compressor  friction  remains  constant.  Of  the  two  methods,  the  clearance 
pocket  or  partial  by-pass  of  the  cylinder  is  probably  the  more  efficient 
for  general  use. 

Another  method  of  operation  is  the  automatic  starting  and  stopping  of 
the  refrigerating  machine,  with  the  automatic  control  designed  to  function 
as  the  load  varies.  This,  however,  is  not  considered  good  practice  as 
mechanical  troubles  develop  and  the  life  of  the  system  is  impaired.  If 
the  equipment  is  kept  in  good  condition,  however,  the  machine  will 
operate  at  maximum  efficiency  so  long  as  it  runs.  The  frequent  starting 
and  stopping  of  large  compressors  is  liable  to  cause  the  power  factor  to 
decrease  if  adequate  allowance  is  not  made. 

All  of  the  methods  described  are  used  from  time  to  time. 

The  methods  of  varying  the  output  of  a  refrigeration  system  which  have 
been  outlined  apply  to  the  reciprocating  type  of  compressor,  although 
variations  in  the  speed  of  the  compressor  to  change  the  refrigerating 
output  are  common  to  all  types  of  mechanical  refrigeration. 

There  is  a  further  method  of  controlling  the  compressor  output  which  is 
particularly  adaptable  to  the  centrifugal  type  of  machine.  This  is  accom- 
plished by  varying  the  amount  of  condensing  water  used  with  the  fluctu- 
ation in  demand  load.  Because  of  the  characteristics  of  the  centrifugal 

172 


CHAPTER  10 — COOLING  METHODS 


type  of  apparatus,  as  the  condensing  water  quantity  is  reduced  and  the 
condensing  temperature  consequently  raised,  the  discharge  pressure  of 
the  centrifugal  machine  rises  correspondingly  and  the  horsepower  input 
to  the  machine  falls  off.  While  this  reduces  the  total  power  input  to  the 
machine,  it  does  not  necessarily  reduce  the  power  input  per  ton  of  re- 
frigeration developed,  as  the  power  input  does  not  drop  with  a  rising  dis- 
charge pressure  as  fast  as  the  refrigerating  effect  produced  drops.  It  is  a 
method,  however,  which  shows  marked  economies  over  the  method 
generally  used  by  the  operating  engineer,  which  is  to  lower  the  suction 
pressure  in  order  to  reduce  the  refrigerating  output  of  the  system. 

Steam  Jet  System 

So  far  the  discussion  has  been  confined  to  reciprocating,  centrifugal,  and 
rotary  compressors.  The  steam  jet  type  of  compressor,  under  certain 
circumstances,  is  desirable  for  use  in  air  conditioning.  Fig.  3  shows  a 
complete  flow  diagram  of  the  system.  The  power  used  for  compressing 


"^a         EVAPORATOR 
CHILLED  VUTER  DISCHARGE 

FIG.  3,    DIAGRAM  OF  STEAM  JET  REFRIGERATION  UNIT 

the  refrigerant  is  steam,  taken  directly  from  the  boiler,  thus  eliminating 
the  mechanical  losses  of  manufacturing  electric  current.  As  the  compres- 
sion ratio  between  the  evaporator  and  condenser  under  normal  circum- 
stances is  large,  the  mechanical  efficiencies  of  the  equipment  are  somewhat 
lower  than  those  of  the  positive  mechanical  type  of  compressor ;  also  the 
condensing  water  requirements  are  considerably  greater,  as  both  the 
refrigerant  and  the  impelling  steam  must  be  condensed. 

The  steam  jet  system  functions  on  the  principle  that  water  under  high 
vacuum  will  vaporize  at  low  temperatures,  and  steam  ejectors  of  the  type 
commonly  used  in  power  plants  for  various  processes  will  produce  the 
necessary  low  absolute  pressure  to  cause  evaporation  of  the  water. 

Fig.  3  shows  a  typical  water  cooling  application.  The  water  to  be 
cooled  enters  the  evaporator  and  is  cooled  to  a  temperature  corresponding 
to  the  vacuum  maintained.  Because  of  the  high  vacuum,  a  small  amount 
of  the  water  introduced  in  the  evaporator  is  flashed  into  steam,  and  as 
this  requires  heat  and  the  only  source  of  heat  is  the  rest  of  the  water  in 
the  evaporator  tank,  this  other  water  is  almost  instantly  cooled  to  a 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

temperature  corresponding  to  the  boiling  point,  determined  by  the 
vacuum  maintained.  The  amount  of  water  flashed  into  steam  is  a  small 
percentage  of  the  total  water  circulated  through  the  evaporator,  amount- 
ing to  approximately  11  Ib  per  hour  per  ton  of  refrigeration  developed. 
The  remainder  of  the  water  at  the  desired  low  temperature  is  pumped  out 
of  the  evaporator  and  used  at  the  point  where  it  is  required. 

The  ejector  compresses  the  vapor  which  has  been  flashed  into  the 
evaporator,  plus  any  entrained  air  taken  out  of  the  water  circulated,  to  a 
somewhat  higher  absolute  pressure,  and  the  vapor  and  air  mix  with  the 
impelling  steam  on  the  discharge  side  of  the  jet.  The  total  mixture  of 
entrained  air,  evaporated  water,  and  impelling  steam  is  discharged  into  a 
surface  condenser  at  a  pressure  which  permits  the  available  condensing 
medium  to  condense  it.  The  resulting  condensate  is  removed  from  the 
condenser  by  a  small  pump,  from  which  it  can  be  discharged  to  the  sewer 
or  returned  to  the  system  in  the  form  of  make-up  water,  or  part  of  it  may 
be  returned  to  the  boiler  feed  pump. 

As  the  normal  temperature  of  water  required  for  air  conditioning 
purposes  is  between  40  F  and  50  F,  with  an  average  temperature  of 
approximately  45  F,  this  type  of  water  cooling  is  particularly  desirable, 
as  the  efficiencies  and  operating  costs  compare  very  favorably  with  other 
types  of  refrigerating  equipment,  especially  in  view  of  the  fact  that  the 
cooling  apparatus  is,  as  a  general  rule,  less  expensive  to  install. 

Approximately  three  times  as  much  condenser  water  is  required  for  the 
steam  jet  cooling  system  as  would  be  necessary  with  other  types  of 
mechanical  refrigeration,  but  as  the  system  can  be  designed  with  a  large 
number  of  jets,  each  of  which  can  be  cut  off  as  the  load  falls  below  maxi- 
mum, constant  refrigerating  efficiency  is  maintained  and  frictional  losses 
and  volumetric  inefficiencies  are  kept  at  a  minimum. 

The  slight  amount  of  air  which  may  be  entrained  in  the  cooled  water  is 
removed  by  a  small  secondary  ejector  which  raises  the  pressure  sufficiently 
so  that  the  air  can  be  discharged  to  the  atmosphere.  A  small  secondary 
condenser,  of  course,  is  necessary  to  condense  the  steam  used  in  the 
secondary  jet. 

Steam  jet  refrigeration  has  an  advantage  where  cooling  towers  are  used 
for  supplying  the  condensing  liquid,  as  there  is  a  great  saving  in  the 
amount  of  steam  used  per  ton  of  refrigeration.  As  the  outdoor  weather 
conditions  vary  the  load  on  the  cooling  system,  the  compression  ratio 
between  the  condenser  and  evaporator  can  be  reduced  and  less  propelling 
steam  need  be  used  per  ton  of  refrigeration  developed.  Roughly,  in  air 
conditioning  work,  mechanical  compressors  show  a  falling  off  of  30  to  40 
per  cent  in  the  power  input  when  using  the  most  economical  arrangement 
of  compressors,  as  the  load  varies  from  100  per  cent  to  25  per  cent  of  the 
rated  capacity;  whereas  with  steam  jet  cooling  equipment,  the  amount  of 
steam  required  for  producing  the  necessary  refrigerating  effect  falls  off  in 
direct  proportion  to  the  load  on  the  system.  When  steam  refrigeration  is  em- 
ployed with  cooling  towers,  the  efficiency  increases  as  the  output  is  reduced. 

Compressors  and  Refrigerants 

There  are  many  different  types  of  compressors,  a  number  of  refrigerants, 
different  types  of  evaporators,  condensers  and  arrangements  of  the  cycle, 
type  has  its  particular  place  and 

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CHAPTER  10 — COOLING  METHODS 


The  generally  used  compressors  are  of  the  following  types: 

1.  Reciprocating  compressors. 

2.  Centrifugal  compressors. 

3.  Rotary  compressors. 

4.  Steam  jet  compressors. 

Over-all  efficiency  of  the  compressor  in  smaller  commercial  installations 
is  not  as  important  a  requirement  as  that  the  whole  unit  require  little 
attention  and  make  a  minimum  of  noise.  The  noise  level  when  the  fan, 
sprays,  and  compressor  are  in  full  operation  should  not  exceed  25  decibels. 
High  compressor  efficiency  appears  as  an  important  factor  only  in  the 
larger  industrial  air  conditioning  systems. 

The  refrigerants  in  most  general  use  in  commercial  and  industrial  air 
conditioning  are  here  listed  in  the  order  of  their  inoffensive  odor  charac- 
teristics : 

1.  Water  vapor. 

2.  Carbon  dioxide. 

3.  Dichlorodifluoromethane. 

4.  Dichloromethane,  sometimes  called  methylene  chloride. 

5.  Methyl  chloride. 

6.  Ammonia. 

7.  Sulphur  dioxide. 

The-  various  types  of  compressors  bear  varied  relationships  to  the 
refrigerants  used  in  both  commercial  and  industrial  air  conditioning. 
Reciprocating  compressors  are  generally  used  for  any  of  the  refrigerants 
listed  except  water  vapor,  dichloromethane,  or  other  low  pressure  refri- 
gerant, and  they  are  used  in  both  commercial  and  domestic  air  conditioning 
systems.  They  have  been  developed  to  a  point  where  their  efficiency  is 
high  and  their  operation  very  satisfactory.  Relatively  low  speed  opera- 
tion makes  them  desirable  for  general  use  in  large  installations.  They  are 
of  two  types,  vertical  and  horizontal,  either  single  or  double  acting.  The 
horizontal  double-acting  compressor  is  not  generally  used  in  air  condition- 
ing except  when  carbon  dioxide  is  used  as  the  refrigerant  in  the  larger 
industrial  systems.  Vertical,  single-acting,  encased  crank,  reciprocating 
compressors  of  the  uniflow  type  with  valves  in  the  pistons  have  proven 
reliable  and  are  used  in  capacities  from  1  hp  to  more  than  100  hp.  Re- 
ciprocating compressors  can  be  used  with  more  refrigerants  than  other 
types  of  compression  units.  For  instance,  when  carbon  dioxide  is  used  as 
the  refrigerant,  a  reciprocating  compressor  is  required  because  of  the 
extremely  high  pressures  and  the  relatively  high  ratio  of  compression. 

The  production  of  refrigeration  at  temperature  levels  from  25  F  to 
55  F  for  general  air  conditioning  involves  special  types  of  refrigerating 
compressors.  Among  these  are: 

1.  Centrifugal  compressors  using  a  volatile  refrigerant. 

2.  Centrifugal  compressors  using  water  as  a  refrigerant. 

3.  Steam  jet  or  vacuum  systems  using  water  as  a  refrigerant. 

4.  Rotary  compressors  using  a  volatile  refrigerant, 


.first  two  types,  centrifugal  compressors,  using  dichloromethane  or 
water  vapor,  can  theoretically  be  used  with  any  of  the  other  refrigerants, 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

but  the  resulting  loss  in  efficiency  with  the  higher  pressure  gases  limits  the 
centrifugal  compressor  to  the  two  refrigerants  named.  At  the  present 
time  the  centrifugal  compressors  are  limited  to  air  conditioning  systems  of 
about  75  hp  and  more.  Centrifugal  compressors  are  usually  built  in  two 
or  more  stages  where  the  compression  ratio  is  high,  and  their  design 
follows  closely  that  of  any  other  centrifugal  equipment,  such  as  general 
service  pumps  and  fans. 

Steam  jet  compressors  which  have  recently  entered  the  field  are  simple 
and  compact  and,  having  no  moving  parts,  they  produce  practically  no 
vibration  but  are  not  economical  for  water  temperatures  much  below 
40  F  or  where  the  cost  of  generating  steam  is  higher  than  the  cost  of 
operation  with  other  prime  movers. 

Rotary  compressors  are  generally  used  for  methyl  chloride  and  dichloro- 
difluoromethane  because  of  their  relatively  low  pressure  and  compression 
ratios.  These  compressors  find  widest  use  for  fractional  tonnage  duty. 

The  source  of  condensing  water  to  some  extent  governs  the  type  of 
refrigerant  used.  If  condensing  water  is  available  at  temperatures  of  not 
more  than  70  to  75  F  any  of  the  refrigerants  mentioned  can  be  used 
economically,  but  if  the  available  condensing  water  temperature  is  above 
80  F,  carbon  dioxide  becomes  uneconomical  as  its  critical  temperature  is 
approximately  88  F.  A  condensing  water  temperature  over  80  F  makes 
the  power  required  for  compression  high.  All  refrigerants  have  critical 
temperatures  and  pressures  sufficiently  high  so  that  their  efficiency  is  not 
materially  affected  by  the  condensing  water  temperatures,  except  in  so 
far  as  this  temperature  affects  the  compression  ratio.  Steam  jet  cooling 
systems  can  use  water  up  to  85  F,  or  even  slightly  warmer. 

The  applicability  of  the  various  refrigerants  is  interesting.  Carbon 
dioxide  is  limited  by  the  condensing  water  temperature;  the  power  con- 
sumption is  slightly  higher  than  that  of  other  refrigerants;  and  the  pres- 
sures are  three  to  four  times  that  of  ammonia. 

The  condenser  pressures  of  methyl  chloride  and  dichlorodifluromethane 
are  approximately  one-half  that  of  ammonia. 

Ammonia,  probably  the  best  known  refrigerant,  has  the  disadvantage 
of  being  toxic,  and  under  certain  circumstances  explosive,  corrosive,  and 
irritating,  even  in  small  quantities  in  the  atmosphere.  Ammonia  is  used 
exclusively  in  the  larger  indirect  or  brine  cooling  air  conditioning  systems. 

Sulphur  dioxide  is  corrosive  and  irritating  even  in  small  quantities  in 
the  atmosphere  and  it  is  toxic  under  certain  circumstances. 

Dichloromethane  operates  at  pressures  below  that  of  the  atmosphere, 
and  it  is  to  some  extent  toxic. 

Dichlorodifluromethane  under  normal  circumstances  is  non-toxic,  non- 
irritating,  and  non-explosive,  but  under  high  temperatures  it  breaks 
down  into  several  obnoxious,  poisonous  components. 

Methyl  chloride,  under  certain  conditions,  is  explosive  and  slightly 
toxic. 

The  steam  ejector  water  vapor  system  has  none  of  the  disadvantages  of 
toxicity,  explosiveness  and  corrosiveness  encountered  in  the  other  refri- 
gerants, but  the  system  operates  at  less  than  atmospheric  pressure.  This, 
however,  is  not  an  important  factor  as  there  are  no  moving  parts  in  the 
compressor  and  the  possibility  of  inleakage  of  air  is  remote  as  all  of  the 

176 


CHAPTER  10 — COOLING  METHODS 


equipment  can  be  welded  air  and  water  tight.  The  supply  of  water  is 
inexhaustible,  and  as  a  refrigerant,  the  make-up  cost  is  negligible.  The 
same  boiler  equipment  can  be  used  for  heating  in  winter  and  for  cooling 
in  summer. 

Electric  Motors 

The  motors  used  for  driving  compressors  can  be  roughly  classified  in 
three  groups:  synchronous,  multispeed,  or  variable  speed.  Further  infor- 
mation on  motors  may  be  found  in  Chapter  17. 

Coolers 

The  types  of  coolers  used  in  connection  with  air  conditioning  work  fall 
into  three  general  groups.  The  first  is  the  direct  cooling  of  water;  the 
second,  direct  cooling  of  air;  and  the  third,  cooling  of  brine  for  circulation 
in  a  closed  system,  which  can  cool  either  water  or  air.  One  method  of  the 
direct  cooling  of  water  is  to  install  direct  expansion  coils  in  the  spray 
chamber  so  that  the  water  sprayed  into  the  air  comes  in  direct  contact 
with  the  cooling  coils.  Another  common  and  efficient  method  of  cooling 
spray  water  is  to  use  a  Baudelot  type  of  heat  absorber  where  the  water 
flows  over  direct  expansion  coils  at  a  rate  sufficiently  high  to  give  efficient 
heat  transfer  from  water  to  refrigerant. 

Another  type  of  spray  water  cooler  is  the  shell  and  tube  heat  exchanger 
in  which  the  refrigerant  is  expanded  into  a  shell  enclosing  the  tubes 
through  which  the  water  flows.  The  velocity  of  the  water  in  the  tubes 
affects  the  rate  of  heat  transfer,  and  as  the  refrigerant  is  in  the  shell  com- 
pletely surrounding  the  tubes  at  all  times,  good  contact  and  a  high  rate  of 
heat  transfer  are  insured.  The  disadvantage  of  such  a  system  is  that  with 
the  falling  off  of  load  on  the  compressor  the  suction  temperature  or  the 
temperature  in  the  evaporator  drops  and  there  is  a  possibility  of  freezing 
the  water  in  the  tubes,  which,  of  course,  might  split  the  tubes  and  allow 
the  refrigerant  to  escape  into  the  water  passage.  This  danger  can  be 
eliminated  by  automatic  safety  devices. 

Another  system  of  cooling  spray  water  is  to  submerge  coils  in  the  spray 
collecting  tank,  or  in  a  separate  tank  used  for  storage.  The  heat  trans- 
mission through  the  walls  of  the  coils,  however,  is  low  and  a  great  deal 
more  surface  is  required  than  for  any  other  type  of  cooler.  However,  with 
large  storage  tanks  this  type  of  cooling  can  be  utilized  to  advantage. 

When  direct  cooling  of  air  is  employed,  the  refrigerant  is  inside  the  coil 
and  the  air  passes  over  it.  Cooling  depends  upon  convection  and  con- 
duction for  removing  the  heat  from  the  air.  The  type  of  coil  used  can  be 
either  smooth  or  finned,  the  finned  coil  being  more  economical  in  space 
requirement  than  the  smooth  coil.  The  fins,  however,  must  be  far  enough 
apart  so  as  not  to  retain  the  moisture  which  condenses  out  of  the  air. 

The  indirect  cooler,  where  brine  is  cooled  by  the  refrigerant  and  the 
resulting  cold  brine  is  used  to  cool  either  air  or  water,  introduces  several 
other  considerations.  It  is  not  the  most  economical  from  a  power  con- 
sumption standpoint,  as  it  is  necessary  to  cool  the  brine  to  a  temperature 
sufficiently  low  so  that  there  is  an  appreciable  difference  between  the 
average  brine  temperature  and  that  of  the  substance  being  cooled.  This 
requires  that  the  temperature  of  the  refrigerant  must  be  still  lower,  and 
consequently  the  amount  of  power  required  to  produce  a  given  amount  of 

177 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

refrigeration  increases  due  to  the  higher  compression  ratio,  but  there  are 
other  considerations  which  make  such  a  system  desirable.  In  the  first 
place,  where  a  toxic  refrigerant  is  undesirable  or  cannot  be  used,  due  to 
fire  or  other  risks  especially  in  densely  populated  areas,  the  brine  can  be 
cooled  in  an  isolated  room  or  building  and  then  be  circulated  through  the 
air  conditioning  equipment  in  perfect  safety  because  it  is  used  to  cool  the 
water  or  air,  without  any  possibility  of  direct  contact  between  the  air  and 
refrigerant. 

When  an  indirect  system  of  cooling  is  used,  it  will  be  found  that  the  heat 
transfer  rate  of  the  water  cooler  is  considerably  higher  as  a  general  rule 
than  that  of  a  direct  expansion  cooler  for  the  same  requirements.  With 
direct  expansion  interchanges,  it  is  almost  impossible  to  keep  the  entire 
system  flooded  with  liquid,  whereas  with  brine  interchangers  the  cooling 
medium  completely  fills  the  space  of  the  interchanger  and  perfect  contact 
is  insured. 

Ice  may  be  used  for  chilling  water  or  air  for  conditioning  work.  Its 
application  is  limited  because  of  the  cost  of  ice,  although  the  efficiency  of 
cooling  is  higher  than  any  other  water  cooling  system.  The  word  "water 
cooling"  is  used  advisedly  in  that  the  direct  cooling  of  air  by  ice  is,  while 
not  impossible,  rather  impractical.  It  might  be  said  that  ice  coolers  are 
economical  for  systems  requiring  a  maximum  of  20  tons  per  24  hours 
where  the  load  fluctuates  considerably,  and  it  is  possible  to  introduce  ice 
only  as  it  is  required  to  cool  water.  The  most  general  method  of  cooling 
water  with  ice  is  to  spray  the  water  over  the  surface  of  the  ice,  insuring  as 
much  contact  as  possible  and  approximating  the  same  performance  as  the 
Baudelot  type  of  cooler.  Because  of  the  large  fluctuations  in  load  in  the 
air  conditioning  system,  the  higher  cost  of  refrigerating  effect  when  ice  is 
used  is  offset  by  the  fact  that  there  are  no  motor  and  condenser  in- 
efficiencies under  partial  load.  Also,  because  the  cost  of  the  mechanical 
refrigeration  equipment  for  the  small  system  is  so  much  higher  per  unit  of 
effect,  the  fixed  charges  are  small  enough  to  overbalance  the  extra  cost 
of  the  ice. 

Condensers 

Condensers  are  usually  either  the  double  pipe  type  or^the  shell  and  tube 
type.  Shell  and  tube  condensers  are  almost  identical  with  coolers. 
Double  pipe  condensers  are  arranged  so  that  water  passes  through  the 
inner  of  two  concentric  pipes,  and  refrigeration  passes  through  the 
annular  space  in  the  outer  pipe.  Where  possible,  there  should  be  counter 
flow  of  the  refrigerant  and  the  condensing  water  to  maintain  maximum 
temperature  differences. 

The  amount  and  temperature  of  the  condensing  water  determine  the 
condensing  temperature  and  pressure,  and  indirectly  the  power  required 
for  compression.  It  is,  therefore,  necessary  to  strike  a  balance  so  that  the 
quantity  of  water  insures  economical  compressor  operation. 

As  part  of  the  condenser,  or  attached  to  it,  there  must  be  storage  space 
for  liquid  refrigerant.  The  installation  of  all  equipment  should  be  made 
accessible  for  inspection,  repair,  and  cleaning.  Both  the  coolers  and 
condensers  should  have  space  for  pulling  tubes. 

Because  there  is  a  decided  tendency  to  conserve  the  water  in  city  mains 
and  most  large  cities  are  restricting  the  use  of  water,  in  order  to  use  air 

178 


CHAPTER  10 — COOLING  METHODS 


conditioning  systems  and  refrigeration  equipment  it  is  often  necessary  to 
install  cooling  towers.  The  cooling  towers,  unfortunately,  produce  the 
warmest  condensing  water  at  the  time  when  the  load  on  the  system  is 
greatest,  so  that  the  refrigeration  equipment  must  be  designed  to  meet 
not  only  the  maximum  load  at  normal  conditions,  but  also  the  maximum 
load  at  abnormal  condensing  water  temperatures.  If  properly  designed, 
this  makes  little  difference  in  the  efficiency  of  operation  throughout  the 
year  except  at  those  times  when  the  condensing  water  temperature  is 
highest.  As  this  occurs  only  for  5  per  cent  of  the  entire  cooling  period  it 
can  be  disregarded  as  a  factor  in  establishing  yearly  operating  costs. 

The  cooling  tower  has  a  certain  advantage  over  the  use  of  water  from 
the  city  mains  in  that  the  temperature  of  the  condensing  water  varies 
directly  with  the  outdoor  temperature  and,  as  pointed  out,  the  refrigera- 
tion load  also  varies  with  this  temperature.  Certain  economies  are  pos- 
sible when  a  cooling  tower  is  used  which  cannot  be  achieved  by  the  use  of 
condensing  water  from  city  mains,  even  where  the  city  water  temperature 
is  extremely  low.  Normally,  the  lowest  city  water  temperature  met  during 
the  summer  months  is  from  65  to  70  F.  This  temperature  range  takes 
place  for  the  entire  cooling  period,  regardless  of  what  the  outdoor  tempera- 
tures are.  With  the  cooling  tower,  the  temperature  of  the  condensing 
water  may  rise  to  80  or  85  F  under  maximum  conditions,  but  under  less 
than  maximum  conditions  the  temperature  of  the  water  off  the  cooling 
tower  drops  considerably,  and  it  has  been  established  that  50  per  cent  of 
the  time  the  outdoor  wet-bulb  temperature  varies  from  60  to  70  F  and  the 
cooling  tower  water,  therefore,  for  the  same  periods,  varies  from  65  to  75  F, 
When  the  outdoor  wet-bulb  temperature  drops  below  60  F,  which  occurs 
approximately  30  per  cent  of  the  time,  the  condensing  water  temperature 
is  still  lower.  The  cost  of  water  used  for  condensing  is  negligible,  as  the 
only  water  required  is  that  used  to  make  up  the  loss  by  evaporation  in  the 
cooling  tower  itself.  See  also  Chapter  11. 

PROBLEMS  IN  PRACTICE 

I  •  In  a  locality  where  the  electric  power  rate  is  based  on  a  demand  charge,  it 
is  desired  to  install  the  smallest  possible  compressor  motor  which  will  provide 
summer  cooling  for  a  300-seat  restaurant  which  operates  6  hours  per  day  from 

II  a.m.  to  2  p.m.,  and  from  5  p.m.  to  8  p.m.    The  refrigeration  load  at  the  peak 
is  28  tons.    If  the  load  factor  for  both  the  noon  and  evening  meals  is  70  per  cent, 
discuss  the  type  of  equipment  which  would  take  the  greatest  advantage  of  the 
reduced  power  rate  at  low  kilowatt  demand. 

A  storage  system  using  a  chilled  water  storage  tank  would  permit  the  installation  of  a 
refrigeration  system  having  the  smallest  motor. 

For  a  28-ton  system  operating  6  hours  per  day  at  a  70  per  cent  load  factor,  on  the  maxi- 
mum day  the  total  heat  removed  would  be, 

28  tons  X  6  hr  X  0.7  =  117.5  ton-hours  per  day. 
If  a  compressor  were  to  operate  24  hours  at  a  constant  rate,  its  average  capacity  would  be 

24  hours =  4"^ t0ns'  °r  aPProx*mately  5  tons.    If  operated  12  hours  per  day,  the 

compressor  capacity  would  have  to  be  increased  to  10  tons. 

A  water  storage  tank  would  store  the  refrigeration  and  allow  off-peak  operation,  so  a 
smaller  compressor  motor  could  be  used.  However,  the  suction  temperature  at  which  the 
compressor  would  be  operated  would  be  lowered  approximately  5  to  10  F.  This  would 
increase  the  horsepower  per  ton  of  refrigeration,  when  dichlorodiflouromethane  is  used, 
approximately  10  per  cent  for  a  5  F  reduction  and  24  per  cent  for  a  10  F  reduction  in  the 

179 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

suction  temperature.    Rather  than  store  the  water  at  too  cold  a  temperature,  it  would 

be  more  economical  to  install  a  larger  storage  tank  and  use  a  higher  temperature. 

A  5-ton  compressor  running  during  periods  when  there  are  no  customers,  namely,  during 

the  15  hours  from  8  p.m,  to  11  a.m.  will  have  stored  15  hr  X  5  tons,  or  75  ton-hours,  of 

refrigeration  in  the  storage  tank  by  11  a.m.    As  one  ton-hour  equals  12,000  Btu,  75  X 

12,000  Btu,  or  900,000  Btu,  will  have  been  stored. 

If  the  apparatus  dew-point  temperature  is  54  F,  and  the  chilled  water  is  supplied  to  the 

air  washer  at  48  F,  it  will  leave  at  54  F.    If  the  water  in  the  storage  tank  is  at  40  F,  the 

temperature  difference  between  the  stored  water  and  the  water  entering  the  wrasher  will 

be  48  F  -  40  F  =  8  F.    This  is  equivalent  to  an  available  8  Btu  of  cooling  effect  per  Ib 

onn  ono 
of  water  stored.     Therefore,  5 —  or  112,500  Ib  of  water  must  be  stored.     This  is 


112,500 


8 
=  13,500  gal  water  to  be  stored,  which  equals 


13,500 


1800  cu  ft  of  water. 


The  storage  tank  to  hold  this  water  might  be  6  ft  high,  7  J^  ft  wide,  and  40  ft  long.  Should 
this  volume  prove  impractical,  a  proportionately  smaller  tank  could  be  used  if  the  water 
storage  temperature  were  reduced.  Should  a  10-ton  refrigeration  system  be  used,  the 
water  quantities  and  tank  capacity  could  be  reduced  by  one  half,  and  the  refrigeration 
plant  need  not  be  started  until  8  a.m.  daily,  which  might  prove  of  additional  advantage. 
If  refrigeration  is  stored  by  freezing  ice  on  coils,  considerable  storage  space  will  be  saved 
but  more  power  input  per  Btu  of  cooling  will  be  required. 

2  •  For  condensing  purposes,  an  air  conditioning  system  uses  city  water  which 
has  an  average  70  F  supply  temperature.  The  following  tahle  lists  the  number 
of  hours  per  year  during  which  definite  wet-bulb  temperatures  and  corre- 
sponding refrigeration  rates  pertain. 


Wet-Bulb 
Temperature 
F 

No.  of 
Hours 
per  Year 

Refrigeration 
Required 
Tons 

80 

6 

284 

79  -  75 

100 

233 

74  -  70 

277 

183 

69  -  65 

330 

157 

64-60 

277 

144 

59  -  55 

158 

79 

54  -  50 

52 

37 

Total  1200  hours 


If  the  power  requirements  of  a  dichlorodifluoromethane  refrigeration  system 
are  in  accordance  with  the  following  data  on  partial  load  operation,  determine 
the  seasonal  power  cost  at  2  cents  per  kwhr: 

284      233       183       157       144        79        37 


Tons  of  Refrigeration 
Kw  per  ton 

Seasonal  power  cost : 


0.89     0.89     0.87     0.86     0.86     0.93     0.97 


WET-BULB 
TEMPERATURE 
P 

TON-HOURS 

KWHR 

80 
79  -  75 
74  -  70 
69  -  65 
64  -  60 
59  -  55 
54  -  50 

Totals 

6  X  284 
100  X  233 
277  X  183 
330  X  157 
277  X  144 
158  X    79 
52  X    37 

«     1,704 
=  23,300 
=  50,700 
=  51,800 
=  39,900 
=  12,500 
=     1,920 

1,704  X  0.89 
23,300  X  0.89 
50,700  X  0.87 
51,800  X  0.86 
39,900  X  0.86 
12,500  X  0.93 
1,920  X  0.97 

=     1,517 
=  20,750 
«  44,100 
«  44,500 
=  34,300 
=  11,600 
=     1,860 

181,824  ton-hours 

158,627  kwhr 

180 

CHAPTER  10 — COOLING  METHODS 


The  158,627  kwhr  at  2  cents  per  kwhr  will  cost  S3,  173. 

158,627  kwhr 

The  average  consumption  will  be  ,Q1  Q0.    -  r  -  =  0.8/3  kw  per  ton. 

i.oifOA'z  ton-nours 

3  •  Using  the  data  from  Question  2,  if  city  water  costs  20  cents  per  thousand 
gallons,  and  if  1.25  gallons  are  used  per  minute  per  ton,  estimate  the  annual 
\vater  cost. 

60  X  1.25  =  75  gal  per  ton-hour. 

181,824  ton-hours  X  75  =  13,620,000  gal  per  year. 

13,620,000X80.20 

-----  —  i7\nn  --  ~~  —  = 
lUuU 


^  ,          r 

,  the  yearly  cooling  water  cost. 


4  •  Using  the  data  of  Question  2,  if  a  cooling  tower  were  installed  for  re-using 
the  condensing  water,  estimate  the  annual  operating  cost  of  a  dichlorodifluoro- 
m  ethane  refrigeration  system  if  the  final  temperatures  of  the  water  leaving  the 
cooling  tower  and  the  kilowatt  input  per  ton  are  the  following  : 

Tons  284       233       183       157       144        79         37 

Temperature  of  water 

leaving  tower,  F  86.7     81.8      76.5      72.1      66.4     61.3      55.6 

Kw  input  per  ton  1.10     0.94     0.85     0.80     0.74     0.59      0.62 


WET-BULB 
TEMPERATURE 
F 

TON-HOURS 

Kw  PER  TON 

1 

KttHR 

80 

1,704 

X 

1.10 

= 

1,875 

79  -  75 

23,300 

X 

0.94 

= 

21,900 

74  -  70 

50,700 

X 

0.85 

= 

43,300 

69  -  65 

51,800 

X 

0.80 

= 

41,400 

64-60 

39,900 

X 

0.74 

= 

29,500 

59  -  55 

12,500 

X 

0.59 

= 

7,370 

54-50 

1,920 

X 

0.62 

= 

1,200 

Totals 


181,824  ton-hours 


146,545  kwhr 


The  146,545  kwhr  at  2  cents  per  kwhr  will  cost  $2,931. 

~  ..         .„  .          146,545  kwhr 

The  average  consumption  will  be  101  00.  r = 

fe  ^  181,824  ton  hours 


0.805  kw  per  ton. 


5  •  If  a  steam  ejector  system  were  used  to  secure  the  refrigeration  for  the  air 
conditioning  system  of  Question  2,  compute  the  annual  steam  cost  if  steam  is 
sold  for  53  cents  per  thousand  pounds  and  if  there  is  an  average  steam  con- 
sumption of  20  Ib  of  steam  per  hour  per  ton  when  used  with  a  cooling  tower 
system. 

181,824  tons  X  20  Ib  of  steam  per  ton  =  3,636,480  Ib  of  steam. 
The  3,636,480  Ib  at  53  cents  per  thousand  pounds  will  cost  $1,929. 

6  •  From  the  data  given  in  the  following  tahle  covering  auxiliary  equipment, 
make  a  comparison  between  the  operating  costs  of  the  complete  dichlorodi- 
fluorome thane  system  of  Question  4  and  the  complete  steam  ejector  cooling 
system  of  Question  5.     A  cooling  tower  is  used  for  condenser  water  recovery. 


Plant  Operation 

Dichlorodifluoromethane 
System 

Steam  Ejector 
System 

Hours  of  operation. 

1200 

1200 

Cooling  tower  fan,  hhp 

17.8 

35.6 

Cooling  tower  pump,  bhp  

30.2 

47.8 

Chilled  water,  gpm 

1200 

1200 

Discharge  head  on  chilled  water 

SyfttftTM^    ft 

75 

75 

Pump  efficiency,  per  cent  

75 

75 

Motor  efficiency,  per  cent 

80 

80 

Chilled  water  temperature,  F  

46 

46 

The  flash  tank  or  evaporator  of  the  steam  ejector  system  is  of  the  open  type, 
the  flash  water  being  pumped  directly  to  the  sprays  of  the  washer  used  for 
cooling  the  air. 

181 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

Dichlorodifluoromethane  System: 
Power  requirements, 

Cooling  tower  fan  17.8  bhp 

Cooling  tower  pump  30.2 

Total  48.0  bhp 


The  water  cooler  in  a  dichlorodifluoromethane  system  of  the  surface  type  requires  no 
additional  pumping  head  other  than  the  friction  drop  through  the  cooler,  which  in  this 
problem  is  estimated  to  be  10  ft.  The  total  pumping  head  is,  therefore,  75  -f  10  =  So  ft. 
Power  required  for  the  chilled  water  system  will  be, 

1200  gpm  X  8.34  Ib  per  gallon  X  85  ft  head 

33,000  ft  Ib  X  0.75  pump  efficiency  P' 

34.3  bhp  X  0.746  X  1200  hr 

-  n  00  -  ;  -  ^—  :  -  =  08,000  kwnr. 
0.80  motor  efficiency 

Thus,  the  total  power  required  by  the  auxiliary  equipment  will  be 
53,700  +  38,300  =  92,000  kwhr. 

The  92,000  kwhr  at  2  cents  per  kwhr  will  cost  $1,840 

The  power  cost  of  refrigeration,  from  Question  4,  is  2t931 

The  total  annual  power  cost,  using  a  dichlorodifluoromethane  system,  is        $4,771 
Steam  Ejector  System: 
Power  requirements, 

Cooling  tower  fan  35,6  bhp 

Cooling  tower  pump  47.8 

Total  83.4  bhp 

T>          f  r      .  ^  83.4  bhp  X  0.746  X  1200  hr        no  OAn  ,     , 

Power  for  cooling  tower  systems  =  -  —  £  -  -  —  *-&-.  -  =  93,300  kwhr. 

0.80  motor  efficiency 

Iii  the  flash  tank  or  water  cooler  of  the  steam  ejector  system,  the  water  is  at  a  pressure 
corresponding  to  the  chilled  water  temperature  required.  In  this  case  it  is  at  46  F,  which 
corresponds  to  an  absolute  pressure  of  0,1532  Ib  per  sq  in.  or  0.3118  in.  Hg.  This  increases 
the  pumping  head  on  the  chilled  water  circulating  pump  by  14.7  —  0.15  =  14.55  Ib  per 
square  inch,  or  33.5  ft.  The  total  pumping  head  is,  therefore,  75.0  +  33.5  =  108.5  ft. 

1200  gpm  X  8.34  Ib  per  gallon  X  108.5  ft  head 

33,000  ft-lb  X  0.75  pump  efficiency  "  P' 

43.7  bhp  X  0.746  X  1200  hr       AQ  Qnn  .     , 

-  ~-x^  -  -  -  -SE~~'  -  =  48,800  kwhr. 
0.80  motor  efficiency 

The  total  power  required  by  the  auxiliary  equipment  is 

93,300  +  48,800  =  142,100  kwhr. 

The  142,100  kwhr  at  2  cents  per  kwhr  will  cost  $2,842 

The  cost  of  the  steam,  from  Question  5,  is  1,929 


The  total  annual  power  cost,  using  a  steam  ejector  system,  is  $4,771 

These  calculations  indicate  that  for  the  assumptions  made,  both  the  dichlorodifluoro- 
methane system  and  the  steam  ejector  system  would  cost  2.6  cents  per  ton-hour  to 
operate.  In  order  to  obtain  a  complete  analysis  it  would  be  necessary  to  compare  the 
fixed  charges  which  include  interest,  depreciation,  obsolescence,  and  maintenance. 
These  are  customarily  computed  at  15  per  cent  of  the  initial  cost  per  annum.  Td  this 
cost  must  be  added  the  cost  of  refrigerant  make-up  per  year.  In  the  steam  system  this 
is  negligible,  but  in  the  dichlorodifluoromethane  system  it  may  be  approximated  at 
from  M  to  %  of  the  refrigerant  charge  per  year. 

182 


Chapter  1 1 

HUMIDIFICATION  AND 
DEHUMIDIFICATION 

Air  Washers*  Atmospheric  Water  Cooling  Equipment,  Cooling 
Towers,  Design  Wet-Bulb  Temperature,  Cooling  Ponds,  Natural 
Draft  Deck  Type  Towers,  Mechanical  Draft  Towers,  Winter  Freezing 

T71  QUIPMENT  for  humidifying  and  dehumidifying  is  of  varied  character 
Py  and  its  functions  will  be  discussed  in  this  chapter.  An  air  washer  is 
essentially  a  chamber  in  which  air  is  brought  in  intimate  contact  with 
water,  the  object  being  (a)  to  wash  the  air  or  (5)  to  regulate  the  moisture 
content  of  the  air  and  at  the  same  time  wash  it.  The  air  comes  in  contact 
with  the  water  by  passing  it  through  water  sprays  or  by  passing  it  over 
surfaces  wetted  by  a  continuous  flow  of  water;  hence  the  classification: 
spray,  scrubber,  and  combination  spray  and  scrubber  type  washers. 

A  washer  chamber  may  be  constructed  of  wood,  or  stone,  but  it  is  most 
often  constructed  of  sheet  metaL  The  lower  portion  of  it  is  specially 
designed  as  a  tank  to  receive  the  water  dropping  through  the  chamber  and 
to  serve  as  a  reservoir  from  which  the  water  may  be  recirculated. 

It  is  desirable  that  air  leaving  a  washer  contain  no  water  in  suspension. 
For  this  reason  eliminators  are  provided  at  the  washer  outlet.  These 
may  be  in  the  form  of  plates  or  baffles  upon  which  the  free  moisture  is 
deposited  as  the  air  is  deflected  through  several  changes  from  its  original 
direction  of  flow.  In  some  washer  units  steel  wool  filter  sections  serve 
as  eliminators.  However,  specially  designed  plates  are  used  more  gener- 
ally than  other  devices  because  they  offer  the  least  resistance  to  the  flow 
of  air,  while  still  performing  effectively  the  function  of  free  moisture 
elimination.  They  also  have  the  advantage  of  acting  as  scrubber  surfaces 
when  flooded. 

It  is  essential  to  uniform  performance  in  a  washer,  that  air  enter  evenly 
distributed  over  the  washer  inlet,  To  insure  this,  a  perforated  plate  or 
eliminator  plates  are  installed  at  the  inlet.  Eliminator  plates  are  now 
more  generally  used.  They  serve  a  second  purpose  in  preventing  the 
escape  of  spray  through  the  washer  inlet. 

Water  is  supplied  to  scrubber  type  units  through  flooding  nozzles.  The 
capacity  of  these  nozzles  varies  with  the  manufacturer  although  a  fair 
value  of  5  gpm  may  be  used.  The  nozzles  are  spaced  on  one-foot  centers 
across  the  top  of  the  washer  over  the  scrubber  plates. 

Water  is  supplied  to  spray  type  units  through  atomizing  nozzles  gener- 
ally arranged  in  banks  across  the  washer.  The  nozzles  spray  either  in  the 
direction  of  the  air  flow,  that  is,  downstream,  or  against  the  air  flow,  or 
upstream.  Nozzle  capacities  vary  with  the  manufacturer,  from  1-J^  to 
2  gpm  at  a  water  pressure  of  about  25  Ib  per  square  inch  which  pressure 

183 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

is  required  for  effective  atomization.  The  spacing  of  spray  nozzles  is 
determined  by  the  water  requirements  of  the  particular  installation.  A 
spray  type  washer  may  contain  one,  two  or  three  banks  of  nozzles  depend- 
ing upon  its  application. 

When  an  air  washer  is  used  for  cleaning  air  it  removes  impurities  and 
dusts.  In  general  it  does  not  function  as  efficiently  in  this  service  as  a 
filter.  For  non-microscopic  soluble  dust  its  efficiency  averages  about 
50  per  cent,  unless  the  concentration  of  dust  is  high.  Its  effectiveness  in 
removing  greasy  microscopic  dust  is  practically  negligible  as  is  also  its 
deodorizing  ability. 

When  a  washer  is  used  to  regulate  the  moisture  content  of  air  it  adds 
moisture  to  (humidifies)  or  removes  moisture  from  (dehumidifies)  the 
air  to  achieve  the  desired  moisture  content.  (See  also  Chapter  3.) 

When  air  passes  through  a  washer  wherein  water  is  circulated  without 
the  addition  or  removal  of  heat,  the  air  tends  to  become  saturated  at  its 
entering  wet-bulb  temperature.  What  occurs  here  is  partial  or  complete 
adiabatic  saturation.  The  total  heat  content  of  the  air  is  unchanged, 
inasmuch  as  the  dry-bulb  temperature  of  the  air  drops  in  proportion  to 
the  amount  of  additional  water  evaporated.  This  action  is  also  known  as 
evaporative  cooling.  A  measure  of  the  washer's  effectiveness  under  these 
conditions  is  its  saturating  efficiency  which  is  equal  to  the  drop  in  dry- 
bulb  temperature  in  per  cent  of  the  entering  wet-bulb  depression.  Other 
things  being  equal,  the  saturating  efficiency  of  a  spray  type  washer  is  a 
function  of  the  number  of  spray  banks  and  the  direction  in  which  they 
spray.  The  following  table  gives  a  general  comparison: 

3  banks — 2  upstream — 1  downstream.™ 100%  saturation  efficiency 

2  banks — 2  upstream 95%  saturation  efficiency 

2  banks — 1  upstream — 1  downstream 85%  saturation  efficiency 

1  bank  — upstream 80%  saturation  efficiency 

1  bank  — downstream 65%  saturation  efficiency 

When  air  passes  through  a  washer  wherein  the  circulated  water  is 
either  cooled  or  heated  before  being  returned  to  the  spray  chamber,  a 
heat  interchange  between  the  air  and  water  occurs,  and  the  air  tends  to 
become  saturated  at  the  temperature  of  the  leaving  water.  The  extent 
to  which  the  leaving  air  and  leaving  water  temperatures  approach  each 
other  is  an  index  to  the  effectiveness  of  the  washer  under  the  operating 
conditions.  The  total  heat  absorbed  by  the  water  in  the  process  equals 
the  total  heat  given  up  by  the  air,  or  the  heat  given  up  by  the  water  equals 
the  heat  absorbed  by  the  air.  Depending  on  whether  the  moisture  con- 
tent of  the  air  is  increased  or  decreased  during  the  operation,  humidifi- 
cation  or  dehumidification  occurs.  Heat  will  be  added  to  or  removed 
from  the  air  as  the  water  supplied  is  of  a  higher  or  a  lower  temperature 
than  the  wet-bulb  temperature  of  the  entering  air. 

For  dehumidifiers  the  ratio  of  the  difference  between  the  leaving  wet- 
bulb  and  the  leaving  water  .to  the  difference  between  the  entering  wet- 
bulb  and  the  entering  ^ater  may  be  figured  as  follows : 

3  banks — 1  downstream — 2  upstream... 0 

2  banks — 2  upstream 5 

2  banks — 1  upstream — 1  downstream. 15 

1  bank  — upstream 20 

1  bank  — downstream 35 

184 


CHAPTER  11 — HUMIDIFICATION  AND  DEHUMIDIFICATION 

Humidifiers  may  be  figured  on  the  same  basis  as  dehumidifiers ;  the 
leaving  water  temperature,  of  course,  will  be  higher  than  the  wet-bulb 
temperature  of  the  leaving  air. 

The  problem  of  cooling  or  heating  the  circulated  water  before  returning 
it  to  the  washer  chamber  is  external  to  the  unit.  It  will  suffice  here  to 
note  that  heating  is  generally  accomplished  by  passing  the  water  through 
closed  hot  water  heaters  or  by  injecting  steam  into  the  water  circuit; 
cooling,  by  passing  the  water  through  closed  coolers  or  over  refrigerating 
coils  in  a  Baudelot  chamber.  Often  in  a  cooling  and  dehumidifiying 
application,  the  refrigerating  coils  are  located  within  the  washer  chamber. 


SPRAY  MANIFOLD 
DRAIN  &  OVERT  LOW 


MANIFOLD 
"  DRAIN  &OVCRFLOW 


FIG.  1.   TYPICAL  SINGLE  BANK  AIR  WASHER       FIG.  2.   TYPICAL  Two  BANK  AIR  WASHER 


Washers  are  sometimes  arranged  in  two  or  more  stages  to  cool  through 
long  ranges  or  to  increase  the  over-all  efficiency  of  heat  transfer  between 
air  and  the  cooling  or  heating  medium  (water,  brine,  etc.) .  A  multi-stage 
washer  is  equivalent  to  a  number  of  washers  in  series  arrangement.  Each 
stage  is  in  effect  a  separate  washer. 

Usually  the  catalog  capacity  of  a  washer  is  expressed  in  cubic  feet  of 
air  per  minute  and  is  based  upon  an  air  velocity  of  500  feet  per  minute 
through  the  gross  cross-sectional  area  of  the  unit  above  the  water  level  in 
its  tank.  At  this  rating  spray  type  washers  handle  about  2-%  gpm  of 
water  per  bank  per  square  foot  of  area,  that  is,  about  5  gpm  per  bank  per 
1000  cfm.  These  proportions  of  air,  water,  area,  and  velocity  may  be 
departed  from  to  meet  the  needs  of  some  particular  job,  but  certain 
limiting  relationships  should  be  observed.  Two  of  the  more  important 
items  are: 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


a.  Choose  a  washer  for  air  velocities  above  approximately  300  fpm  and  below 
approximately  600  fpm.     Velocities  outside  this  range  are  likely  to  result  in  faulty 
elimination  of  entrained  moisture. 

b.  When  a  high  saturating  efficiency  is  required,  select  a  two  or  three  bank  spray 
type  unit,  having  a  total  water  capacity  of  not  less  than  15  gpm  per  100  cfm. 

The  area  of  a  washer  may  be  dictated  by  space  limitations  outside  the 
washer,  such  as  headroom,  or  by  space  requirements  inside  washer,  such 
as  face  area  needed  by  a  bank  of  cooling  coils.  The  length  of  a  washer  is 
determined  by  the  number  of  spray  banks,  or  scrubber  plates,  and  if 
cooling  coils  are  installed  in  the  unit,  by  the  number  of  banks  of  coils. 
Roughly,  a  spray  space  of  about  2  ft  6  in.  in  length  is  required  for  each 
bank  of  sprays,  (the  leaving  eliminators  require  about  1  ft  6  in.,  entering 
eliminators  about  1  ft). 

The  resistance  to  air  flow  through  an  air  washer  varies  with  the  type 
eliminators,  number  of  banks  of  sprays,  direction  of  spray,  type  of  scrub- 


DISTRIBUTING 
THERMOMETER 


DIAPHRAGM  VALVE 


FIG.  3.    AIR  WASHER  WITH  SPRAY  WATER  HEATING  ARRANGEMENT 

ber  plates,  and,  if  cooling  coils  are  located  in  unit,  by  their  size  and  type. 
Washers  should  be  selected  to  limit  static  resistances  below  0.50  in. 

Power  Requirements 

The  approximate  power  requirement  for  passing  10,000  cfm  of  ^ air 
through  a  humidifier  of  the  spray  type  by  a  fan  of  78  per  cent  mechanical 
efficiency  is  given  in  Table  1,  this  being  the  fan  brake  horsepower  for 
various  velocities  and  static  pressure  losses.  Allowance  should  be  made 
for  variations  in  static  pressure  due  to  the  use  of  different  diffuser  plates 
or  inlet  louvers  and  for  variations  in  fan  efficiencies. 

ATMOSPHERIC  WATER  COOLING  EQUIPMENT 

To  successfully  operate  a  refrigerating  plant  or  a  condensing  turbine, 
the  heat  from  the  compressed  refrigerant  or  the  discharged  steam  must  be 
removed  and  dissipated.  This  is  accomplished  ordinarily  by  first  trans- 
ferring the  heat  of  the  gas  to  water  in  a  heat  exchanger.  If  the  plant  is 
situated  on  the  banks  of  a  river  or  lake,  an  intake  may  be  had  upstream  or 
at  a  considerable  distance  from  the  discharge,  to  prevent  mixing  of  the 
heated  discharged  water  with  the  inlet  water.  If  the  source  of  water  is  a 
city  supply  or  well  water,  the  discharge  water  may  be  run  into  the  nearest 
sewer  or  open  waterway.  Lacking  an  unlimited  water  supply,  or  in  cases 
where  city  water  is  too  expensive  or  where  the  water  available  contains 

186 


CHAPTER  11 — HUMIDIFICATION  AND  DEHUMIDIFICATION 


dissolved  salts  which  would  quickly  form  scales  on  the  heat-exchanging 
apparatus,  it  is  necessary  to  recirculate  the  water,  and  to  cool  it  after  each 
passage  through  the  heat-exchanger  by  exposure  to  air  in  an  atmos- 
pheric water  cooling  apparatus. 

Air  has  a  capacity  for  absorbing  heat  from  water  when  the  wet-bulb 
temperature  of  the  air  is  lower  than  the  temperature  of  the  water  with 
which  it  is  in  contact.  The  rapidity  with  which  this  transfer  of  heat  occurs 
depends  upon  (1)  the  area  of  water  in  contact  with  the  air,  (2)  the  relative 
velocity  of  the  air  and  water,  and  (3)  the  difference  between  the  wet-bulb 
temperature  of  the  air  and  the  temperature  of  the  water.  Because  the 
changes  in  rate  do  not  occur  in  direct  proportion  to  changes  in  the  govern- 
ing factors,  data  on  the  performance  of  atmospheric  water  cooling  equip- 
ment are  largely  empirical. 

TABLE  1.     APPROXIMATE  FAN  BRAKE  HORSEPOWER 

Requirements  for  passing  10,000  cfm  of  air  through  humidifiers  at  various  velocities  and  static  pressures. 
Mechanical  efficiency  of  fan — 78  per  cent. 


30  DEG  ELIMINATORS  SPACED 

45  DEG  ELIMINATORS  SPACED 

VELOCITY 

ON  1-Ys 

IN.  CENTERS 

ON  2-}4  IN.  CENTERS 

Static  Pressure 

! 

Static  Pressure 

In.  Water 

In.  Water 

500 

0.20 

I           0.40 

0.40 

0.80 

550 

0.24 

|           0.48 

0.48 

0.97 

600              !            0.29 

i           0.58 

0.58 

1.15 

650                          0.34 

•    i           0.68 

0.68 

1.35 

As  the  heat  content  of  the  air  increases,  its  wet-bulb  temperature  rises. 
(See  Chapter  1.)  Because  it  is  impractical  to  leave  the  air  in  contact 
with  water  for  a  long  enough  time  to  permit  the  wet-bulb  temperature  of 
the  air  and  the  temperature  of  the  water  to  reach  equilibrium,  atmos- 
pheric water  cooling  equipment  aims  to  circulate  only  enough  air  to  cool 
the  water  to  the  desired  temperature  with  the  least  possible  expenditure 
of  power. 

Cooling  Towers 

In  an  air  washer,  humidifier  or  dehumidifier,  the  air  is  first  conditioned 
by  water  to  change  its  moisture  and  temperature,  and  it  is  then  sent  to 
the  place  where  it  is  to  be  used.  In  water  cooling  equipment  the  tem- 
perature of  the  water  is  reduced  by  air,  and  the  cooled  water  is  carried  to 
its  point  of  usage.  In  the  air  washer,  an  excess  of  water  is  used  to  con- 
dition a  fixed  quantity  of  air,  while  in  water  cooling  equipment,  an  excess 
quantity  of  air  is  used  to  cool  a  fixed  quantity  of  water. 

Both* types  of  equipment  have  a  common  basis  of  design,  however,  in 
that  the  size  of  the  equipment  is  determined  by  the  quantity  of  air  that 
must  be  handled.  With  the  air  washer,  the  size  of  the  equipment  is  fixed 
by  the  quantity  of  air  to  be  conditioned,  and  the  amount  of  conditioning 
is  controlled  by  the  quantity  and  temperature  of  the  water  supplied  and 
its  method  of  application.  With  water  cooling  apparatus,  its  size  and  the 
quantity  of  air  required  bear  no  direct  relation  to  the  quantity  of  water 
being  cooled,  but  vary  through  a  wide  range  for  different  services  and 
conditions. 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


Sizes  of  Equipment 

Assuming  a  definite  quantity  of  water  to  be  cooled,  the  size  and  design 
of  atmospheric  cooling  equipment  are  affected  by  the  following  factors: 
L    Temperature  range  through  which  the  water  must  be  cooled. 

2.  Number  of  degrees  above  the  wet-bulb  temperature  of  the  entering  air  to  which 
the  water  temperature  must  be  reduced. 

3.  Temperature  of  the  atmospheric  wet-bulb  at  which  the  required  cooling  must  be 
performed. 

4.  Time  of  contact  of  the  air  with  the  water.    (This  involves  height  or  length  of  the 
apparatus  and  velocity  of  air.) 

5.  Surface  of  water  exposed  to  each  unit  quantity  of  air. 

6.  Relative  velocity  of  air  and  water. 

TABLE  2.    CONDENSER  DESIGN  DATA 


GAS 

MAXIMUM  PRESSURE 
DESIRED  IN 
CONDENSER 

GAS  TEMPERATURE 
IN  CONDENSEH 

F 

LEAVING  HOT  WATER  TEMPERATURE 
F 

Best  Design 

Average  Design 

Steam 

28  in.  vacuum  

99.7 
114.3 
126.0 

96.0 
86.0 
100.0 
100.0 

97 
110 
120 

92 
83 
96 
96 

93 
105 
114 

88 
81 
92 
93 

Steam 

27  in.  vacuum  

Steam 

26  in.  vacuum  

Ammonia  

185  Ib  gage 
head  pressure  
1030  Ib  gage 
head  pressure  
102  Ib  gage 
head  pressure  
1171bgage 
head  pressure  

Carbon  dioxide.. 

Methyl^ 
chloride 

Dichlorodi- 
fluoromethane 

Items  1,  2,  and  3  are  established  by  the  type  of  service  and  geographical 
location,  while  items  4,  5,  and  6  depend  upon  the  design  of  the  equipment. 
The  establishment  of  a  proper  cooling  range  depends  upon : 

1.  Type  of  service  (refrigerating,  internal  combustion  engine  and  steam  condensing). 

2.  Wet-bulb  temperature  at  which  the  equipment  must  operate  satisfactorily. 

3.  Type  of  condenser  or  heat-exchanger  used. 

Because  the  design  of  an  entire  plant  is  usually  affected  by  the  quantity 
and  temperature  of  the  cooling  water  supply,  plants  should  be  designed 
for  cooling  water  conditions  which  can  be  most  efficiently  attained.  The 
first  consideration  is  usually  the  limiting  temperature  of  the  plant.  For 
example,  if  an  ammonia  compressor  refrigerating  plant  is  to  be  designed 
for  185  Ib  head  pressure  as  a  normal  maximum,  the  limiting  temperature 
of  the  ammonia  in  the  condenser  is  96  F.  Should  the  ammonia  temperature 
go  above  this  figure  the  head  pressure  will  exceed  185  Ib  and  power  con- 
sumption increases.  To  obtain  this  head  pressure,  the  temperature  of  the 
circulating  water  leaving  the  condenser  must  always  be  less  than  96  F 
by  an  amount  depending  upon  the  size  and  design  of  the  condenser,  the 
quantity  of  water  being  circulated,  and  the  refrigerating  tonnage  being 
produced.  A  condenser  having  a  large  surface  per  ton  of  refrigeration 
may  be  designed  to  operate  satisfactorily  with  the  leaving  hot  water 
temperature  within  3  deg  or  4  deg  of  the  ammonia  temperature  cor- 
responding to  the  head  pressure,  while  a  small  condenser  might  require 
a  10  deg  difference. 

188 


CHAPTER  11 — HUMIDIFICATION  AND  DEHUMIDIFICATION 

Table  2  lists  several  gases  with  data  as  to  the  temperatures  and  pres- 
sures for  which  commercial  condensers  are  designed.  Internal  combustion 
engines  have  limiting  hot  water  temperatures  of  125  F  to  140  F.  The 
cooling  of  such  fluids  as  milk  or  wort  has  variable  requirements  and  is 
usually  done  in  counter-flow  heat-exchangers  in  which  the  leaving  circu- 
lating water  is  at  a  much  higher  temperature  than  is  the  leaving  fluid. 

The  temperature  range,  once  the  hot  water  temperature  is  approxi- 
mately known,  depends  upon: 

1.  Maximum  wet-bulb  temperature  at  which  the  full  quantity  of  heat  must  be 
dissipated. 

2.  Efficiency  of  the  atmospheric  cooling  equipment  considered. 

Design  Wet- Bulb  Temperatures 

The  maximum  wet-bulb  temperature  at  which  the  full  quantity  of 
water  must  be  cooled  through  the  entire  range  is  never,  in  commercial 
design,  the  maximum  wet-bulb  temperature  ever  known  to  exist  at  the 
location  nor  the  average  wet-bulb  temperature  over  any  period.  The 
former  basis  would  require  atmospheric  cooling  equipment  several  times 
greater  than  normal  size,  and  the  latter  would  result  during  a  large  part  of 
the  time,  in  higher  condenser  water  temperatures  than  those  for  which  the 
plant  was  designed.  For  instance,  the  maximum  wet-bulb  temperature 
recorded  in  New  York  City  is  88  F,  and  the  July  noon  average  for  64 
years  is  close  to  68  F.  Yet  in  the  years  1925  to  1931,  inclusive,  there  were 
but  6  hrs  per  year  when  the  wet-bulb  temperature  reached  80  F  or  more, 
and  there  were  975  hours  in  the  average  summer  (June  to  September, 
inclusive)  when  the  wet-bulb  temperature  was  68  F  or  above.  As  these 
975  hours  represent  a  third  of  the  summer  period,  cooling  equipment 
based  upon  the  noon  average  July  wet-bulb  of  68  F  would  be  inadequate. 
Commercial  practice  is  to  choose  a  wet-bulb  temperature  for  refrigeration 
design  purposes  which  is  not  exceeded  during  more  than  5  to  8  per  cent 
of  the  summer  hours  (75  F  for  New  York  City),  with  somewhat  lower 
requirements  for  steam  turbines  and  internal  combustion  engines.  This 
difference  is  made  because  the  heaviest  load  on  a  refrigerating  plant  is 
coincident  with  high  wet-bulb  temperatures,  whereas  the  heaviest  electric 
power  demand  occurs  either  in  the  winter  or  after  nightfall  in  summer, 
when  the  wet-bulb  temperature  is  low.  Table  1,  Chapter  8,  shows  safe 
design  wet-bulb  temperatures  which  will  not  be  exceeded  more  than  8  per 
cent  of  the  time  in  an  average  summer. 

Knowing  the  hot  water  temperature  and  the  wet-bulb  temperature  for 
which  the  equipment  must  be  designed,  the  cold  water  temperature  must 
be  chosen  to  place  the  requirement  within  the  efficiency  range  of  the  type 
of  atmospheric  water  cooling  apparatus  to  be  used.  Efficiency  of  atmos- 
pheric water  cooling  apparatus  is  expressed  as  the  percentage  ratio  of  the 
actual  cooling  range  to  the  possible  cooling  range.  Since  the  wet-bulb 
temperature  of  the  entering  air  is  the  lowest  temperature  to  which  the 
water  could  possibly  be  cooled  this  is : 

Percentage  cooling  efficiency  of  atmospheric  water  cooling  equipment  = 

(hot  water  temperature  —  cold  water  temperature  )  X  100 
hot  water  temperature  —  wet-bulb  temperature  of  entering  air 

189 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


Efficiencies  of  various  types  of  atmospheric  water  cooling  apparatus 
vary  through  wide  limits,  depending  upon  air  velocity,  concentration  of 
water  per  square  foot  of  area,  and  the  type  of  equipment.  The  commercial 
range  of  efficiencies  is  given  in  Table  3  although  unusual  designs  may 
operate  outside  these  ranges. 

From  consideration  of  the  factors  which  include  the  cooling  range  and 
design  wet-bulb  temperature,  the  quantity  of  water  required  can  be 
calculated  from  the  amount  of  heat  to  be  dissipated.  The  normal  amounts 
of  heat  to  be  removed  from  various  parts  of  the  cooling  equipment  are: 

Compressor  refrigeration 220  to    270  Btu  per  minute  per  ton 

Condenser  turbine 950  to    980  Btu  per  pound  of  steam 

Steam  jet  refrigerating  appartus 1030  to  1150  Btu  per  pound  of  steam 

Diesel  engine 2800  to  4500  Btu  per  horsepower 

Cooling  Ponds 

A  natural  pond  is  often  used  as  a  source  of  condensing  water.  The 
hot  water  should  be  discharged  close  to  the  surface  at  the  shore  line,  as 
natural  air  movement  over  the  surface  of  the  water  will  cause  evaporation 

TABLE  3.    EFFICIENCY  OF  ATMOSPHERIC  WATER  COOLING  EQUIPMENT 


EQUIPMENT 


COOLING  EFFICIENCY— PEH  CENT 


MjnJ.mi.TTn. 

Usual 

Maximum 

Spray  Ponds 

30 

45  to  55 

60 

Spray  Towers                   

40 

45  to  55 

60 

Natural  Draft  Deck  or  Atmospheric 
Towers 

35 

50  to  70 

90 

Mechanical  Draft 

35 

55  to  75 

90 

and  carry  away  heat.  Because  increased  density  due  to  the  loss  of  heat 
causes  the  cooled  water  to  sink  to  the  bottom  of  the  pond,  the  suction 
connection  for  intake  water  should  be  placed  as  far  below  the  surface  as 
possible,  and  at  as  great  a  distance  from  the  discharge  as  practicable. 

Spray  Cooling  Ponds 

* 

The  spray  pond  consists  of  a  basin,  above  which  nozzles  are  located  to 
spray  water  up  into  the  air.  Properly  designed  spray  nozzles  break  up  the 
water  into  small  drops,  but  not  into  a  mist  because  the  individual  drops 
must  be  heavy  enough  to  fall  back  into  the  basin  and  not  drift  off.  The 
water  surface  exposed  to  the  air  for  cooling  is  the  combined  area  of  all  the 
small  drops.  Since  the  rate  of  heat  removal  by  atmospheric  water  cooling 
is  a  function  of  the  area  of  water  exposed  to  the  air,  the  difference  in 
temperature  between  the  water  and  the  wet-bulb  temperature  of  the  air, 
the  relative  velocity  of  air  and  water,  and  the  duration  of  contact  of  the 
air  with  the  water,  a  much  larger  quantity  of  heat  may  be  dissipated  in  a 
given  area  with  the  spray  pond  than  with  the  cooling  pond,  because  of  (1) 
the  speed  with  which  the  drops  travel  as  they  are  propelled  into  the  air 
and  fall  back  into  the  water  basin,  (2)  the  increased  wind  velocity  at  a 
point  above  the  surrounding  structures  or  terrain,  (3)  the  increased 

190 


CHAPTER  1 1 — HUMIDIFICATJON  AND  DEHUMIDIFICATION 

volume  of  air  used,  and  (4)  the  vastly  increased  area  of  contact  between 
air  and  water. 

Spray  pond  efficiencies  are  increased  by  (1)  elevating  the  nozzles  to  a 
higher  point  above  the  surface  of  the  water  in  the  basin,  (2)  increasing  the 
spacing  between  nozzles  of  any  one  capacity,  (3)  using  smaller  capacity 
nozzles,  to  decrease  the  concentration  of  water  per  unit  area,  and  (4) 
using  smaller  nozzles  and  increasing  the  pressure  to  maintain  the  same 
concentration  of  water  per  unit  area.  Usual  practice  is  to  locate  the 
nozzles  from  3  ft  to  6  ft  above  the  edge  of  the  basin,  to  supply  from  5  Ib  to 
12  Ib  pressure  at  the  nozzles,  using  nozzles  spraying  from  20  gpm  to 
60  gpm  each  and  spacing  them  so  the  average  water  delivered  to  the 
surface  of  the  pond  is  from  0.1  gpm  per  square  foot  in  a  small  pond  to 
0.8  gpm  per  square  foot  in  a  large  pond. 

Increasing  the  pressure,  spacing  the  nozzles  farther  apart,  or  increasing 
the  elevation  of  the  nozzles  will  increase  the  cross-section  of  spray  cloud 
exposed  to  the  air,  and  therefore  increase  the  quantity  of  air  coming  in 
contact  with  the  water.  Best  results  are  obtained  by  placing  the  nozzles 
in  a  long  relatively  narrow  area  located  broadside  to  the  wind. 

Spray  ponds  may  be  located  on  the  ground  if  they  have  an  earthen  or 
a  concrete  basin,  or  they  may  be  placed  on  roofs  having  special  waterproof 
roofing.  To  prevent  excessive  drift  loss,  or  the  carrying  of  entrained 
water  beyond  the  edge  of  the  pond  by  the  air  on  the  leeward  side,  louver 
fences  are  required  for  roof  locations  and  for  those  ground  locations  where 
space  is  so  restricted  that  the  outer  nozzles  cannot  be  located  at  least 
20  ft  to  25  ft  from  the  edge  of  the  basin.  Such  fences  usually  are  con- 
structed of  horizontal  louvers  overlapping  so  the  air  is  forced  to  turn  a 
corner  in  passing  through  the  fence,  and  the  heavier  drops  of  water  are 
thrown  back,  owing  to  their  inertia.  The  louvers  also  restrict  the  flow  of 
air,  particularly  at  the  higher  wind  velocities,  and  thus  further  reduce  the 
possibility  of  water  being  carried  off.  The  height  of  an  effective  fence 
should  be  equal  to  the  height  of  the  spray  cloud.  Louver  boards  are 
preferably  of  red  gulf  cypress  or  California  redwood  supported  on  cast- 
iron,  steel  or  wood  posts,  Where  building  ordinances  forbid  the  use  of 
combustible  materials,  sheet  metal  is  customarily  used. 

Algae  formations  may  be  a  considerable  nuisance  in  a  spray  pond. 
Such  growths  are  killed  by  the  periodic  addition  of  potassium  permanga- 
nate to  the  pond  water.  Addition  of  the  dissolved  chemical  should  be 
made  until  the  water  holds  a  faint  pink  color  for  at  least  15  min. 

Spray  Cooling  Towers 

Where  not  more  than  30,000  Btu  per  minute  are  to  be  dissipated,  the 
spray  cooling  tower  is  a  satisfactory  apparatus.  The  word  tower  in  this 
connection  is  somewhat  of  a  misnomer  as  the  apparatus  is  essentially  a 
narrow  spVay  pond  with  a  high  louver  fence.  As  usually  built,  the  nozzles 
spray  down  from  the  top  of  the  structure  and  the  distance  from  the  center 
of  the  nozzle  system  to  the  fence  on  either  side  is  not  more  than  half  the 
distance  that  the  nozzles  are  elevated  above  the  water  basin.  Heights 
range  from  6  ft  to  15  ft  and  the  total  width  of  a  structure  is  not  usually 
greater  than  its  height.  Spray  cooling  towers  occupy  less  space  on  small 
jobs  than  spray  ponds  of  equivalent  capacities  because  the  towers  have 
a  capacity  of  from  0.6  gpm  to  1.5  gpm  per  square  foot  of  tower  area.  The 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

louvers  are  continually  wet,  and  so  add  to  the  surface  of  water  exposed 
to  the  cooling  air. 

Natural  Draft  Deck  Type  Towers 

In  past  years  most  of  the  atmospheric  water  cooling  on  refrigeration 
work  has  been  done  with  natural  draft  deck  type  towers,  which  are  also 
referred  to  as  wind  or  atmospheric  towers.  These  towers  consist  of  heavy 
wooden  or  steel  framework  from  15  ft  to  80  ft  high  and  from  6  ft  to  30  ft 
wide,  having  open  horizontal  lattice-work  platforms  or  decks  at  regular 
intervals  from  top  to  bottom,  and  a  catch  basin  at  the  foot.  The  hot 
water  is  distributed  over  the  upper  part  of  the  structure  by  means  of 
troughs,  splash  heads,  or  nozzles,  and  it  drips  from  deck  to  deck  down  to 
the  basin.  The  object  of  the  decks  is  to  arrest  the  fall  of  the  water  so  as  to 
present  efficient  cooling  surfaces  to  the  air,  which  passes  through  the 
tower  parallel  to  the  decks.  The  decks  also  add  to  the  area  of  water 
surface  exposed  to  the  air,  but  since  they  furnish  a  resistance  to  air  flow, 
too  many  decks  are  a  detriment. 

To  prevent  the  loss  of  water  on  the  leeward  side  of  the  tower,  wide 
splash  boards  are  attached  at  regular  intervals  from  top  to  bottom.  These 
boards  or  louvers  extend  outward  and  upward,  and  in  most  designs  the 
top  edge  of  each  louver  extends  above  the  bottom  edge  of  the  one  above  it. 

Efficiency  of  a  deck  tower  is  improved,  within  limits,  by  increased 
height,  increased  length,  or  increased  width,  The  first  two  increase  the 
area  of  water  exposed  to  the  wind,  and  the  latter  increases  the  time  of 
contact  of  the  air  with  the  water. 

Wind  Velocities  on  Natural  Draft  Equipment 

Since  natural  air  movement  is  the  prime  requirement  for  a  deck  type 
tower,  spray  cooling  tower,  or  spray  pond,  the  apparatus  must  be  de- 
signed to  produce  the  desired  cooling  on  days  when  the  wind  velocity  is 
below  average  when  the  wet-bulb  temperature  is  at  the  maximum  chosen 
for  design,  and  when  the  plant  is  operating  at  full  load.  The  apparatus 
must  also,  for  best  results,  be  located  with  its  longest  axis  at  right  angles 
to  the  direction  of  the  prevailing  hot  weather  breeze.  Table  1  Chapter  8, 
gives  the  average  summer  wind  velocities  and  directions  in  representative 
cities.  Natural  draft  cooling  equipment  should  be  designed  to  operate 
properly  with  not  more  than  one-half  of  the  average  wind  velocity,  and  in 
no  case  should  it  need  a  wind  velocity  of  more  than  5  mph.  It  is  obvious 
that  natural  draft  towers  and  other  natural  draft  equipment  must  be  so 
located  that  they  are  not  obstructed  by  trees,  buildings,  or  other  wind 
deflectors. 

Mechanical  Draft  Towers 

Mechanical  draft  towers  usually  consist  of  vertical  shells,  constructed 
of  wood,  metal,  or  masonry,  in  which  water  is  distributed  uniformly  at  the 
top  and  falls  to  a  collecting  basin  at  the  bottom.  The  inside  of  the  tower 
may  be  filled  with  wood  checker-work  over  which  the  water  drips,  or  the 
water  surface  may  be  presented  to  the  air  by  filling  the  entire  inside  of  the 
structure  with  spray  from  nozzles.  Air  is  circulated  through  the  tower 
from  bottom  to  top  by  forced  or  induced  draft  fans.  Since  the  air  flows 
counter  to  the  water,  the  air  is  in  contact  with  the  hottest  of  the  water 

192 


CHAPTER  11 — HUMIDIFICATION  AND  DEHUMIDIFICATION 

just  before  leaving  the  top  of  the  tower,  and  each  unit  of  air  picks  up  more 
heat  than  a  similar  unit  would  on  natural  draft  equipment,  so  the  me- 
chanical draft  tower  cools  water  by  using  less  air  than  the  other  types  of 
equipment  need.  As  movement  of  the  air  through  the  towers  is  obtained 
by  power-consuming  fans,  it  is  essential  that  the  air  used  be  reduced  to  a 
minimum  so  as  to  secure  the  lowest  possible  operating  cost. 

The  efficiency  of  a  mechanical  draft  tower  is  increased  by  increasing 
height,  area,  or  air  quantity.  Increasing  the  height  increases  the  length 
of  time  the  air  is  in  contact  with  the  water  without  affecting  seriously  the 
fan  power  required,  but  it  increases  the  pumping  power  needed.  In- 
creasing the  area  while  maintaining  constant  fan  power  increases  the  air 
quantity  somewhat  and  because  of  louvered  velocities  it  increases  the 
time  this  air  is  in  contact  with  the  water.  The  surface  area  of  water  in 
contact  with  the  air  is  increased  in  both  cases.  Increasing  the  air  quantity 
decreases  the  time  the  air  is  in  contact  with  the  water,  but,  since  a  greater 
quantity  is  passing  through,  the  average  differential  between  the  water 
temperature  and  the  wet-bulb  temperature  of  the  air  is  increased,  and 
this  speeds  up  the  heat  transfer  rate.  Increased  air  quantities  are 
obtained  only  at  the  expense  of  increased  fan  power,  which  increases 
approximately  as  the  cube  of  the  air  quantity.  Air  velocities  through 
mechanical  draft  towers  vary  from  250  f pm  to  600  f pm  over  the  gross  area 
of  the  structure. 

Mechanical  draft  water  cooling  equipment  may  be  set  up  inside  build- 
ings, where  it  usually  draws  its  air  supply  from  the  general  space  in  which 
it  is  installed,  and  discharges  its  exhaust  air  through  a  duct  to  the  outside. 
Indoor  cooling  towers  may  be  either  of  the  wood-filled  or  the  spray-filled 
type.  In  many  cases  where  little  height  but  considerable  area  is  available, 
water  is  cooled  in  a  spray-filled  structure  similar  to  an  air  washer,  with 
the  air  passing  horizontally  through  the  apparatus  and  being  discharged 
through  a  duct  to  the  outside.  Such  apparatus  does  not  have  the  counter 
flow  advantage  of  the  vertical  mechanical  draft  water  cooling  equipment, 
and  therefore  requires  a  much  larger  excess  of  air  for  proper  operation. 
Air  velocities  and  operating  powers  are  considerably  above  those  required 
by  vertical  mechanical  draft  water  cooling  equipment. 

Make-up  Water 

Since  the  atmospheric  water  cooling  equipment  performs  its  functions 
chiefly  by  evaporating  a  portion  of  the  water  in  order  to  cool  the  re- 
mainder, there  is  a  continual  drain  on  the  quantity  of  water  in  the  system, 
and  this  loss  must  be  replaced.  Approximately  1  gal  of  water  is  lost  for 
every  1000  gal  of  water  cooled  per  degree  of  cooling  range;  so  if  1000  gpm 
of  water  are  cooled  through  a  10  deg  range,  10  gpm  of  water  will  be  re- 
quired to  replace  evaporated  water.  Replacement  supply  is  usually 
regulated  by  a  float  control  valve.  Because  the  evaporation  of  the  water 
leaves  behind  the  salts  which  the  water  contained,  high  concentration  of 
salts  may  make  chemical  treatment  of  the  make-up  water  necessary  to 
avoid  excessive  deposits  in  the  condensers. 

Winter  Freezing 

If  atmospheric  water  cooling  equipment  is  operated  in  freezing  weather, 
the  water  may  be  cooled  below  freezing  temperature  so  ice  forms  and 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

collects  until  its  weight  causes  damage.  To  obviate  freezing  during  con- 
tinued operation,  the  efficiency  of  the  apparatus  may  be  lowered.  This 
is  done  on  the  spray  pond  and  the  spray  cooling  tower  by  reducing  the 
quantity  of  water  fed  to  the  apparatus,  thereby  lowering  the  pressure  at 
the  nozzles  and  increasing  the  size  of  the  drops  produced.  On  the  deck 


TABLE  4.  COMPARISON  OF  VARIOUS  TYPES  OF  ATMOSPHERIC  WATER  COOLING  EQUIPMENT 

Figures  indicate  order  of  desirability 


COOLING 
POND 

SPRAY 
POND 

SPRAY 
TOWER 

DECK 
TOWER 

MECHANICAL 
DRAFT 

INDOOR 
TOWEH 

Cost                           

X 

2 

1 

3 

4 

5 

Area 

5 

4 

3 

2 

1 

X 

Height                           

1 

2 

3 

4-5 

4-5 

X 

Weight  per  sq  ft 

X 

X 

1 

3 

4 

2 

Independence  of  wind  velocitv 

6 

3 

4 

5 

1-2 

1-2 

Drift  nuisance 

1 

6 

5 

4 

2-3 

2-3 

Make-up  water  required 

1 

6 

5 

4 

2-3 

2-3 

Pumping  head                            

1 

2 

3 

4-5 

4-5 

6 

Maintenance                  .         

2 

1 

3 

4 

5 

6 

Suitability  for  congested  districts  

X 

5 

4 

3 

1 

2 

Water  quantity  required  for  definite 
result 

6 

5 

4 

1-2 

1-2 

3 

*Not  comparable. 


tower  the  upper  system  may  be  shut  off  and  a  secondary  distribution 
system  put  in  service  midway  down  the  height  of  the  tower.  The  water 
will  be  kept  above  freezing  because  it  will  have  shorter  contact  with  the 
air.  The  mechanical  draft  tower  can  be  protected  by  reducing  the  air 
flow  through  the  tower,  by  stopping  or  reducing  the  speed  of  the  fans,  or 
by  partially  closing  dampers. 

If  the  system  is  operated  intermittently  in  freezing  weather,  water  in 
the  basin  may  freeze  and  the  expansion  of  the  ice  may  do  harm.  Freezing 
during  intermittent  operation  can  be  prevented  only  by  draining  the 
water  basin  when  it  is  out  of  service.  On  small  roof  installations,  a  tank 
large  enough  to  hold  all  the  water  in  the  system  is  often  installed  inside 
the  building  and  the  basin  is  drained  into  this  by  gravity,  the  pump  suc- 
tion being  taken  from  this  inside  tank. 

A  comparison  of  various  types  of  water  cooling  equipment  is  given  in 
Table  4. 

PROBLEMS  IN  PRACTICE 

1  •  What  three  systems  of  humidification  are  used  in  textile,  printing,  and 
lithographic  plants? 

a.  Indirect:    Introduction  of  moistened  air  into  the  rooms. 

b.  Direct:    Spraying  of  moisture  into  the  rooms. 

c.  Combined:    Direct  and  indirect  as  above. 

2  •  How  may  relative  humidity  be  controlled? 

a.  If  constant  room  temperature  is  to  be  maintained: 

1.  To  maintain  a  constant  relative  humidity,  the  dew  point  must  be  kept  constant. 

194 


CHAPTER  11 — HUMIDIFICATION  AND  DEHUMIDIFICATION 

2.  To  increase  the  relative  humidity,  the  dew  point  must  be  raised. 

3.  To  decrease  the  relative  humidity,  the  dew  point  must  be  lowered. 

b.  If  constant  dew  point  is  to  be  maintained: 

1.  To  maintain  a  constant  relative  humidity,  the  room  temperature  must  remain 
constant. 

2.  To  increase  the  relative  humidity,  the  room  temperature  must  be  lowered. 

3.  To  decrease  the  relative  humidity,  the  room  temperature  must  be  raised. 

c.  With  varying  dew-point  temperatures: 

1.  To  maintain  a  constant  relative  humidity,  the  room  temperature  must  vary 
directly  and  in  almost  equal  amount  with  the  dew  point. 

2.  To  increase  the  relative  humidity,  the  difference  between  room  temperature  and 
dew  point  must  be  decreased. 

3.  To  decrease  the  relative  humidity,  the  difference  between  room  temperature  and 
dew  point  must  be  increased. 

d.  With  varying  room  temperatures: 

1.  To  maintain  a  constant  relative  humidity,  the  dew  point  must  vary  directly  and  in 
almost  equal  amount  with  the  room  temperature. 

2.  To  increase  the  relative  humidity,  the  difference  between  dew  point  and  room 
temperature  must  be  decreased. 

3.  To  decrease  the  relative  humidity,  the  difference  between  dewr  point  and  room 
temperature  must  be  increased. 

3  •  In  industrial  air  conditioning  plants,  what  are  the  four  sources  of  heat 
which  must  be  taken  into  consideration  in  the  design  of  a  system? 

a.  Heat  transfer  from  the  outside  air. 

b.  Body  heat  from  employees. 

c.  Sun  effect. 

d.  Heat  equivalent  of  power  consumed  in  driving  machinery,  in  lighting,  and  in  manu- 
facturing processes  in  general. 

4  •  Why  do  cooling  towers  give  best  results  when  the  humidity  of  the  air  is  low? 

The  cooling  of  water  by  dropping  it  through  air  depends  mostly  upon  the  evaporation  of 
the  water.  If  the  relative  humidity  of  the  air  is  low,  the  water  vapor  will  be  readily 
absorbed  and  carried  away,  while  if  the  humidity  of  the  air  is  high,  its  capacity  to  pick 
up  water  vapor  is  less  and  the  water  is  cooled  less  with  the  same  exposure  to  air. 

5  •  What  performance  tests  should  be  given  air  washers? 

a.  Capacity. 

b.  Resistance. 

c.  Visible  entrainment  of  free  moisture. 

d.  Humidifying  efficiency. 

e.  Cleaning  effect. 

6  •  What  are  the  several  different  types  of  water-cooling  towers? 

a.  Those  with  forced  draft. 

b.  Those  with  natural  draft  open  to  the  atmosphere. 

c.  Those  with  natural  draft  closed  to  the  atmosphere. 

d.  Those  with  combined  natural  and  forced  draft. 

7  •  What  are  the  different  types  of  air  washers? 

a.  Spray,     b.  Wet  scrubber,     c.  Combination  spray  and  scrubber. 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

8  •  What  is  the  saturation   efficiency  for  an  air   washer  with   the  common 
variations  in  spray  arrangement? 

For  three  banks,  two  up-stream  and  one  down-stream  ..     100C< 

For  two  banks,  both  up-stream - -. 95r^ 

For  two  banks,  one  up-stream  and  one  down -stream    ,  85  % 

For  one  bank,  up-stream 80^ 

For  one  bank,  down-stream., .     ..  ,,         65^r 

9  •  Upon  what  air  velocity  are  air  washers  usually  rated? 

500  fpm,  through  the  area  above  the  tank. 

10  •  What  wet-bulb  temperature  for  the  outside  air  is  usually  selected  in  air 
conditioning  design  when  cooling  is  to  be  accomplished? 

One  which  is  not  exceeded  more  than  5  to  8  per  cent  of  the  time  in  the  locality  where  the 
plant  is  to  be  situated. 

11  •  Where  should  the  suction  connection  be  placed  in  a  cooling  pond? 

As  far  below  the  surface  as  possible  and  as  far  away  from  the  discharge  as  practicable 

12  •  What  chemical  is  used  to  kill  algae  formations  in  spray  ponds? 

Potassium  permanganate. 

13  •  What  is  the  usual  amount  of  spray  water  delivered  to  a  cooling  pond  per 
square  foot  of  pond  area? 

From  0.1  gpm  on  small  sizes  to  0,8  gpm  on  large  sizes. 

14  •  What  is  the  usual  amount  of  water  delivered  in  cooling  towers  per  squar 
foot  of  area? 

From  0.6  to  1.5  gpm. 

15  •  About  how  much  water  is  lost  by  evaporation  in  atmospheric  cooling? 

About  1  gal  per  1000  gal  for  each  degree  of  cooling  range. 

16  •  How  is  freezing  obviated  in  cooling  pond  sprays? 

The  pressure  and  quantity  of  water  is  lowered  so  that  the  drops  become  of  increased  size 
and  do  not  freeze  so  readily. 

17  •  What  is  the  cause  of  a  high  concentration  of  salts  in  the  cooling  water  of 
an  atmospherically  cooled  system? 

The  constant  evaporation  of  a  small  portion  of  the  water  leaves  salts  behind  to  accumu- 
late in  the  unevaporated  water. 


196 


Chapter  12 

UNIT  AIR  CONDITIONERS  AND 
CONDITIONING  SYSTEMS 

Definition,  Advantages  and  Uses,  Functions,  Sources  of  Refrigera- 
tion and  Heat,  Types  and  Locations,  Construction  of  Apparatus., 
Installation.,  Basis  of  Equipment  Ratings,  Calculation  of  Required 
Capacity,  Approximate  Costs 

A  IR  conditioning  systems  fall  into  two  general  types  known  as  the  unit 
jC\,  type  and  the  central  type.  A  unit  air  conditioner  is  an  assembly  of 
parts,  such  as  fans,  humidifiers,  coils,  controls,  and  other  equipment, 
which  form  a  complete  unit  at  the  point  of  manufacture.  This  usually 
restricts  the  size  of  the  unit  to  a  capacity  below  10,000  cfm.  With  the 
unit  conditioner,  the  performance  is  the  responsibility  of  the  manu- 
facturer. This  is  in  contradistinction  to  a  central  air  conditioning  system 
which  may  produce  the  same  results  but  for  which  the  various  parts  are 
purchased  separately  and  assembled  by  the  contractor  on  the  job,  who 
guarantees  the  performance  of  the  assembled  system. 

Unit  Air  Conditioner 

A  unit  air  conditioner  generally  has  a  capacity  less  than  30,000  Btu  per 
hour  for  cooling,  or  60,000  Btu  per  hour  for  heating,  to  make  it  suitable 
for  the  space  to  be  conditioned.  If  it  does  not  provide  simultaneous 
control  of  at  least  four  of  the  recognized  functions  of  air  conditioning  (see, 
p.  201)  the  apparatus  should  be  classified  as  a  unit  heater  or  unit  venti- 
lator (Chapter  13)  or  as  a  unit  cooler,  a  humidifier,  or  a  window-type 
ventilator. 

The  apparatus,  instead  of  being  wholly  self-contained,  may  depend 
upon  separately  located  parts  piped  to  supply  heating,  cooling,  or  humi- 
difying mediums  to  the  unit.  A  duct  may  supply  outdoor  air  for  circu- 
lation, but  ducts  are  seldom  used  for  air  discharge  and  recirculation. 

When  the  term  unit  conditioner  is  applied  to  such  set-ups  as  the  com- 
bination of  a  filter  and  a  fan  in  a  housing  to  be  used  with  gravity  warm  air 
furnaces,  or  to  humidifiers  and  heating  coils  to  be  used  with  steam  or  hot- 
water  boilers  to  comprise  a  unified  central  air  conditioning  plant,  the 
usage  of  the  term  is  inaccurate;  such  devices  may  be  designated  as 
accessory  units,  but  this  leads  to  confusion.  However,  since  such  accessory 
equipment  is  used,  a  description  and  discussion  of  its  several  types  are 
given  in  the  next  few  paragraphs  before  the  main  topic  of  this  chapter, 
unit  heaters,  is  taken  up. 

197 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

Accessory  Central  Fan  Conditioning  Apparatus 

This  Includes  every  kind  of  equipment  constituting  an  accessory  to  an 
existing  or  new  system  for  warm  air  heating  service,  and  also  certain 
forms  of  conditioning  equipment  used  with  hot  water  or  steam  boilers  in 
residential  service.  Some  of  these  accessories  provide  only  a  fan  and  an 
air  filter,  while  others  include  humidifying  and  cooling  functions.  The 
performance  of  such  equipment  is  influenced  by  the  outside  temperature 
and  humidity;  the  conditions  surrounding  the  house  or  apartment,  such 
as  construction  and  exposure  to  sun ;  the  type  of  heating  system  to  which 
the  apparatus  is  attached;  and  the  location  of  the  device  on  the  heating 
system.  Many  of  these  installations  are  of  limited_capacity  and  effective- 
ness ;  conservative  manufacturers  will  be  discriminating  in  their  claims 
for  added  comfort  from  the  use  of  such  equipment,  depending  on  its 
design  and  functions. 


i  *  Ml     --ll-^K— — .A,r  to  p00m& 


FIG.  1.    FURNACE  ACCESSORY  UNIT 

A  feature  of  the  fan-and-filter  accessory  unit  is  its  availability  ^  for 
ventilation  in  summer;  it  makes  possible  a  rapid  cooling  in  the  evening, 
after  the  outdoor  air  temperature  has  dropped  below  that  of  the  rooms. 
If  the  fan  is  large  enough  completely  to  change  the  air  in  the  building 
served  every  two  or  three  minutes,  the  effect  will  be  similar  to  that  from 
so-called  attic  fans,  (see  Chapter  13),  with  the  important  advantage  that 
the  air  is  filtered.  Fans  of  smaller  capacity,  proportioned  only  for  ^the 
winter  heating  duty,  may  also  provide  an  appreciable  measure  of  cooling. 
Another  advantage  is  improved  headroom  in  the  basements  of  residences, 
obtainable  by  substituting  horizontal  ducts  for  those  of  comparatively 
steep  pitch  necessary  when  gravity  air  circulation  is  depended  upon.  A 
fan-and-filter  accessory  using  a  dry-mat  type  of  filter,  applied  to  a  warm 
air  furnace,  is  shown  in  Fig.  1. 

A  more  elaborate  unit  (Fig.  2),  for  use  with  a  hot  water  heating  boiler 
provides  heating,  humidification,  filtering,  and  positive  air  circulation  in 
winter;  the  heating  coil  may  be  used  also  in  summer  with  mechanical 
refrigeration  or  for  circulating  city  water  or  chilled  water  from  an  ice  tank, 
to  provide  cooling  and  dehumidification.  The  disposition  of  fans,  the 
cloth  filter  of  bag  design,  the  spray  type  humidifier,  as  well  as  noise 

198 


CHAPTER  12 — UNIT  AIR  CONDITIONERS  AND  CONDITIONING  SYSTEMS 


elimination  features  comprising  canvas  collars  at  the  fan  outlets  and 
rubber  pads  under  the  fan  bedplate,  are  indicated  in  Fig.  3.  The  use  of  a 
single  element  for  both  winter  and  summer  functions  tends  to  reduce  the 
first  cost,  although  it  adds  some  complications  in  piping. 

Another  assembly  of  air  conditioning  equipment  with  a  standard 
heating  furnace,  in  this  instance  burning  gas  to  provide  warm  air,  is 
shown  in  Fig.  4.  The  apparatus  comprises  an  air  filter,  a  motor-driven 
fan,  and  an  air  washer.  No  refrigeration  is  used  with  this  equipment. 


Hot  Water 
Better 


.ftir  to  ffooms 


•Fitter 


Fan 


FIG.  2.    UNIT  WITH  HOT  WATER  BOILER 


Rubbe, 
Pads, 


FIG.  3. 


HEATING  AND  COOLING  UNIT 
WITH  CLOTH  FILTER 


Return  tfir  from 


-/?/r  Wisher 
'  Furnace 

FIG.  4.    GAS  FIRED  FURNACE  UNIT 

For  oil  fuel,  the  unit  shown  in  Fig.  5  can  be  installed  to  obtain  filtered, 
warmed,  and  humidified  air.  An  oil  burner  and  a  heat  exchanger  provide 
the  heat.  A  cooling  section  may  be  inserted  between  the  fan  and  the  heat 
exchanger,  cold  water  being  circulated  through  the  cooling  element.  For 
automatic  control,  a  room  thermostat  is  provided  to  start  the  oil  burner 
whenever  the  temperature  falls.  The  rising  temperature  in  the  heat 
exchanger  causes  a  second  thermostat  to  start  the  fan.  As  soon  as  the 
temperature  in  the  house  rises  to  normal,  the  room  thermostat  shuts  down 
the  oil  burner  and  operates  the  thermostat  controlling  the  fan. 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

Having  disposed  of  the  accessory  central  fan  conditioning  apparatus, 
the  balance  of  this  chapter  will  concern  only  the  unit  air  conditioner  as 
defined  in  Chapter  41. 

ADVANTAGES  AND  USES  OF  UNIT  AIR  CONDITIONERS 

Unit  air  conditioners  are  suitable  for  commercial  and  comfort  applica- 
tions because  they  permit  installation  without  seriously  disturbing  the 
building  occupants,  and  they  allow  rearrangement  or  a  change  in  capacity 
to  suit  changed  requirements  occasioned  by  new  tenants.  Tenants  may 
even  furnish  their  own  installations  and  remove  the  apparatus  from  the 
premises  at  the  expiration  of  their  leases.  In  some  types  of  buildings,  the 
installation  costs  are  lower  for  unit  conditioners  than  those  for  central  fan 
systems,  and  costs  are  further  lowered  in  that  there  is  no  need  for  space  in 
which  to  house  a  conditioning  plant.  The  choice  between  unit  and 


To  Room, 


FIG.  5.    OIL  FIRED  UNIT 

central  systems  will,  in  many  instances,  require  a  close  study  of  instal- 
lation conditions  at  the  site,  and  a  preparation  of  comparative  cost 
estimates,  in  addition  to  a  consideration  of  the  more  intangible  factors. 

Industrial  Uses 

The  origin  of  the  unit  conditioner,  like  that  of  air  conditioning  itself, 
was  in  the  industrial  field  for  maintaining  desired  atmospheric  conditions 
in  rooms  or  sections  of  manufacturing  plants  where  structural  limitations 
or  service  requirements  made  a  central  system  uneconomic.  Industrial 
applications  continue  to  offer  an  important  market  for  unit  conditioners, 
in  bakeries,  candy  factories,  drug-manufacturing  plants,  laboratories, 
produce-storage  rooms,  printing  plants,  and  similar  places. 

Commercial  Uses 

The  most  active  field  for  unit  conditioners  at  the  present  time  is  in 
commercial  establishments,  such  as  barber  and  beauty  shops,  funeral 
parlors,  retail  and  specialty  stores,  and  small  restaurants,  where  increased 
patronage  or  larger  purchases  per  customer  offer  economic  justification  of 
first  cost  and  operating  expense. 

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CHAPTER  12 — UNIT  AIR  CONDITIONERS  AND  CONDITIONING  SYSTEMS 

The  air-handling  units  installed  overhead  in  Pullman  cars,  diners  and 
coaches,  in  the  middle  or  at  the  ends  of  the  cars  which  discharge  the  air 
horizontally  at  the  ceiling  level,  are  essentially  unit  conditioners.  How- 
ever, because  of  special  construction  to  meet  space  limitations  and  other 
requirements,  they  are  unsuitable  for  general  use  and  are  not  further 
discussed. 

Personal  Uses 

The  major  recent  development  in  the  air  conditioning  industry  has  been 
in  new  and  improved  types  of  unit  conditioners  suitable  for  apartments, 
homes,  hotel  rooms,  and  offices.  These  uses  demand  apparatus  that  is 
compact,  of  unobtrusive  appearance  and  in  harmony  with  the  room 
finish  and  furnishings,  quiet  in  operation,  automatic,  and  reliable. 

As  with  all  new  major  appliances  for  the  home,  problems  of  relatively 
high  first  cost,  of  comparatively  rapid  obsolescence  and  of  operating 
expense  demand  the  continuous  close  attention  of  manufacturers  and  of 
others  interested  in  developing  the  potential  market.  Unit  conditioners 
are  still  distinctly  in  the  pioneering  stage  where  such  problems  must  be 
met  and  solved  if  development — especially  of  residential  units — is  to 
proceed  as  fast  as  it  should.  Public  understanding  of  residential  air  con- 
ditioning still  requires  cultivation  in  order  to  cBspel  fears  of  possible 
excessive  operating  costs  and  of  possible  high  obsolescence  due  from 
frequent  model  changes.  Progress  in  this  direction  is  being  helped  by  the 
increasing  efforts  of  manufacturing  companies  which  are  now  spending 
large  sums  to  insure  sound  promotion  of  unit  conditioners.  Likewise,  the 
National  Better-Housing  Program  inaugurated  in  1934  is  likely  to  prove  of 
real  value  to  the  air  conditioning  industry  and  to  accelerate  greatly  the 
rate  of  public  acceptance  and  installation  of  unit  conditioners. 

FUNCTIONS  OF  UNIT  CONDITIONERS 

Unit  air  conditioners  may  be  classified  as  the  all-year  unit,  the  summer 
unit,  and  the  winter  unit.  The  all-year  unit  performs  all  of  the  functions 
of  an  air  conditioning  system ;  namely ,  cooling,  dehumidification,  heating, 
humidification,  air  circulation,  air  cleaning — with  or  without  a  supply  of 
fresh  air — and  a  simultaneous  control  of  all  functions.  The  summer  unit 
must  provide  cooling,  dehumidification,  air  circulation,  and  air  cleaning; 
the  winter  unit  must  provide  heating,  humidification,  air  circulation,  and 
air  cleaning.  Either  of  these  seasonal-use  units  may  or  may  not  provide  a 
fresh  air  supply  and  a  simultaneous  control  of  the  functions. 

In  some  instances,  winter-type  units  equipped  with  filters  for  air 
cleaning  and  with  fresh  air  connections  may  be  operated  in  summer  for 
ventilation,  but  the  system  cannot  then  properly  be  said  to  provide 
all  year  conditioning.  It  is  important  that  the  features  and  limitations  of 
the  specific  apparatus  be  carefully  explained  to  a  prospective  user,  so  that 
disappointments  and  complaints  concerning  operating  results  may  be 
avoided. 

The  functions  listed  are  performed  by  the  unit  conditioners  offered  by 
different  manufacturers  in  various  ways,  some  of  which  appear  in  the 
following  outline.  See  the  next  few  pages  for  more  detailed  explanations 
of  cooling  and  heating  theories  and  methods. 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

1.  Cooling: 

By  coils,  usually  of  finned  type,  for  direct  expansion  of  refrigerant  from  a  self- 
contained  unit  or  from  a  remotely-located  compressor. 

By  coils,  of  finned  type,  for  brine  or  cold  water  from  separate  refrigeration  plant, 
or  for  cool  water  from  city  mains,  private  wells,  or  an  ice  water  tank. 
By  water  sprays. 
By  passage  of  air  over  ice  cakes. 

2.  Dehumidification : 

By  lowering  the  air  temperature  below  the  dew  point,  using  any  of  the  devices 

outlined  for  cooling. 

By  adsorption  materials,  such  as  silica  gel  or  activated  alumina. 

3.  Heating: 

By  coils,  usually  of  finned  type,  for  steam  or  hot  water  from  system  distribution 
mains. 

By  electric  heating  elements. 
By  gas  burners. 

4.  Humidification: 

By  evaporating  or  entraining  water  by  an  air  current,  from  wetted  surfaces  or 
water  sprays. 

5.  Air  circulation : 

By  motor-driven  fans  which  discharge  air  into  room  at  points,  in  directions  and 
with  velocities  that  insure  adequate  ventilation  without  drafts;  air  discharge 
usually  through  top,  at  a  slight  angle  from  vertical. 

6.  Air  cleaning: 

By  mechanical  filters. 

By  water  washing  with  sprays. 

By  water  washing  by  contact  with  condensation  or  by  trickling  water  on  cooling 

coils  or  a  mesh  cell. 

7.  Fresh  air  supply: 

By  air  connection  from  outdoors,  usually  through  adjustable  window  ducts  at 
rear  of  housing,  with  mixing  dampers  for  control  of  volume  of  recirculated  room 
air  taken  in  through  louvers  at  each  end. 

8.  Control: 

By  manual  adjustment  or  automatic  regulation,  by  thermostats  or  hygrostats. 

SOURCES  OF  REFRIGERATION 
Mechanical  Refrigeration — Direct  and  Indirect 

In  general,  mechanical  refrigeration  uses  the  low-temperature  evapora- 
tion of  a  liquid  to  absorb  heat  in  a  set  of  coils.  The  resulting  vapor  is 
restored  to  its  original  liquid  state  by  compressing  and  condensing  it, 
abstracting  the  heat  by  passing  water  or  air  over  a  second  set  of  coils  at 
the  outlet  side  of  the  compressor.  Power  for  compression  is  usually 
supplied  by  an  electric  motor.  The  apparatus,  exclusive  of  the  evaporator 
or  cooling  coil,  is  known  as  a  condensing  unit.  Two  methods  are  available 
for  applying  mechanical  refrigeration  to  unit  conditioners. 

The  direct-expansion  system  provides  for  admitting  the  refrigerant 
through  a  pressure  reducing  (expansion)  valve  to  the  cooling  coil,  where 
its  evaporation  causes  chilling  of  the  surface  over  which  the  circulated  air 
passes.  Under  this  method,  the  equipment  cost  is  low,  the  refrigerant 
lines  need  not  be  insulated,  the  apparatus  is  compact,  and  the  operating 
expense  is  minimized  by  the  avoidance  of  heat  leakage  and  by  the  higher 

202 


CHAPTER  12 — UNIT  Am  CONDITIONERS  AND  CONDITIONING  SYSTEMS 

permissible  suction  pressure  at  the  compressor  inlet  (as  compared  with 
the  indirect  system).  However,  because  of  possible  hazards  from  leaks, 
direct  expansion  is  usually  prohibited  in  hospitals  and  places  of  public 
assembly. 

The  indirect-expansion  system  uses  a  water-submerged  coil  in  a  tank 
near  the  condensing  unit,  for  evaporation  of  the  refrigerant.  The  chilled 
water  or  brine  is  then  delivered  under  pressure  by  a  motor-driven  pump 
for  distribution  to  the  cooling  coils  in  the  individual  unit  conditioners, 
returning  again  to  the  tank.  This  avoids  the  possibility  of  refrigerant 
vapors,  whether  toxic  or  not,  leaking  into  the  conditioned  rooms.  Code 


•Fan  Motor 


•— Ha net  Control 
Valve 


Line. 


FIG.  6.    ROOM  COOLING  UNIT 

limitations  on  the  quantity  of  refrigerant  in  the  air  conditioning  apparatus 
are  overcome,  and'a  central  condensing  unit  may  be  made  to  serve  rooms 
on  different  floors  or  in  remote  parts  of  a  building,  without  violating 
safety  regulations.  Difficulties  that  occur  with  compressor  operation  at 
less  than  50  per  cent  of  rated  capacity  are  avoided  through  the  use  of  a 
thermostat  that  shuts  down  the  compressor  when  the  tank  water  tem- 
perature reaches  the  set  minimum;  operation  is  had  at  constant  suction 
pressure,  independent  of  the  number  of  unit  conditioners  running.  With 
proper  choice  of  temperatures  at  which  the  compressor  starts  and  stops 
under  thermostatic  control,  there  is  less  cycling  than  with  the  direct 
expansion  system.  Under  favorable  conditions,  the  cooling  coil  may  be 
supplied  with  steam  or  hot  water  for  winter  heating,  thereby  simplifying 
the  construction  of  the  unit  conditioner,  although  at  the  expense  of  some 
complication  in  valved  connections.  However,  the  cold  water  tank  and 
circulating  pump  take  up  room,  and  the  cost  of  suitably  insulated  dis- 
tribution piping  is  greater  than  that  of  equivalent  liquid  lines  and  suction 
returns  for  a  direct-expansion  system. 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

Separate  Condensing  Units 

The  separate  condensing  unit,  for  mechanical  refrigeration  with  unit 
conditioners  that  are  not  self-contained,  comprises  the  assembly,  on  a 
bedplate,  of  a  compactly  arranged  compressor  with  motor,  drive,  con- 
denser, liquid  receiver,  and  automatic  controls.  The  cylinder  jacket  of 
the  compressor  and  the  condenser  may  be  cooled  with  water  under  pres- 
sure, or  with  air  supplied  by  a  fan  mounted  integrally  with  the  compressor. 
A  condensing  unit  connected  to  a  single  unit  conditioner  is  shown  in 
Fig.  6. 

Steam-Jet  Apparatus 

Stearn-jet  (vacuum)  refrigeration  may  be  used  in  localities  served  by 
district  steam  mains,  or  in  buildings  with  boiler  plants  available  for 
summer  use.  While  avoiding  power-driven  compressors,  the  steam-jet 
apparatus  requires  an  appreciable  amount  of  power  for  auxiliary  pumps, 
and  an  increased  quantity  of  cooling  water  to  absorb  the  heat  from  the 
motive  steam  in  addition  to  that  abstracted  from  the  conditioned  rooms. 
Most  installations  of  this  type  are  of  large  capacity — above  20  tons 
refrigeration — but  recently  developed  equipment  is  available  for  instal- 
lations as  small  as  2  to  5  tons. 

City  or  Well  Water 

Systems  installed  near  the  Great  Lakes  or  in  other  regions  where  low 
cost  cooling  water  is  available  in  summer  may  often  use  this  water 
directly  in  the  coils  of  air  conditioning  units.  In  certain  other  places,  well 
water  can  be  obtained  in  sufficient  quantity  at  moderate  pumping 
expense.  Restrictions  on  bulk  use  of  water,  or  on  discharge  of  large 
volumes  into  the  sanitary  sewers  may  prevent  direct  cooling. 

Ice 

Two  methods  of  using  ice  are  applicable :  direct,  with  air  circulated  by  a 
fan  over  ice  cakes  in  an  insulated  tank  within  the  room  served;  indirect, 
with  an  ice-melting  tank  remote  from  the  unit  conditioners,  circulating 
chilled  water  to  coils  in  the  units  by  means  of  a  motor-driven  pump.  The 
direct  method  has  been  employed  with  portable  room  coolers  for  hotel 
guest  rooms,  hospitals,  and  residences,  where  the  demand  for  air  con- 
ditioning is  moderate  and  variable  with  respect  to  rooms  served  from  day 
to  day,  and  where  it  is  feasible  to  move  the  units  into  a  service  room  or 
kitchen  for  emptying  and  icing.  The  indirect  method  is  identical  with 
that  common  in  theaters  using  ice,  except  that  the  water  after  spraying 
over  the  ice  is  pumped  to  unit  conditioners  instead  of  to  a  central  fan 
system. 

SOURCES  OF  HEAT 
Steam  or  Hot  Water  Coils 

The  heating  coils  of  unit  conditioners  for  all-year  or  winter  service  are 
available  for  either  steam  or  hot  water,  supplied  at  low  or  high  pressure 
from  building  heating  plants.  Because  the  relatively  high  Btu  per  hour 

204 


CHAPTER  12 — UNIT  AIR  CONDITIONERS  AND  CONDITIONING  SYSTEMS 

outputs  for  heating  (usually  1.5  to  3  times  the  rate  for  cooling),  under 
thermostatic  control,  may  produce  disturbances  in  small  heating  systems, 
it  is  usually  necessary  that  two-pipe  steam  systems  operate  at  all  times 
above  atmospheric  pressure,  and  that  hot  water  systems  have  forced 
circulation  with  a  pump.  Unless  the  room  space  occupied  by  radiators 
in  an  existing  building  is  needed  for  the  unit  conditioners  or  for  other 
purposes,  it  is  preferable  that  some  or  all  of  them  be  retained,  so  that  the 
unit  conditioners  need  supply  only  sufficient  heat  to  permit  their  satis- 
factory operation  for  humidification. 

Electric  Elements  or  Gas  Burners 

Where  energy  is  available  at  low  cost,  electric  heating  elements  may  be 
used  in  place  of  steam  or  hot  water  coils  for  winter  service.  Evaporation 
of  water  for  winter  humidification  may  likewise  be  accomplished  electri- 
cally. More  uniform  control  of  temperature  and  humidity  is  practicable 
with  electricity,  because  the  heating  elements  may  be  divided  into  sections 
separately  connected  through  thermostatically  controlled  switches. 
However,  wiring  connections  must  be  larger  than  needed  for  summer 
conditioning  with  a  compressor  built  into  the  unit;  for  instance,  the 
power  for  a  unit  rated  at  24,000  Btu  per  hour  for  winter  heating  is  about 
seven  times  that  used  for  12,000  Btu  per  hour  of  summer  cooling  by  the 
same  unit. 

A  new  unit  conditioner  employing  the  adsorption  method  for  summer 
dehumidification  is  fitted  with  gas  burners  for  winter  heating.  A  part  of 
the  humidification  is  supplied  by  the  water  vapor  resulting  from  com- 
bustion of  hydrogen  in  the  gas  fuel,  and  the  remainder  by  evaporation 
from  a  heated- water  receptacle. 

TYPES  AND  LOCATIONS  OF  UNIT  CONDITIONERS 
Fixed 

The  majority  of  unit  conditioners  are  designed  for  floor  mounting, 
preferably  under  windows.  However,  when  radiators  for  winter  heating 
occupy  the  window  space  and  it  is  not  desired  to  shift  them  or  to  eliminate 
them  by  using  all-year  type  floor  units,  the  location  may  be  against 
interior  partitions  or  alongside  permanently  situated  furniture.  In  all 
cases  care  must  be  taken  to  insure  that  the  direction  of  the  air  discharge 
will  not  cause  drafts  that  may  be  objectionable  to  occupants.  When  out- 
door air  for  ventilation  is  taken  through  the  unit,  the  under-window 
position  is  advantageous,  since  it  permits  using  a  short  inlet  duct  from 
louvers  in  a  filler  panel  permanently  inserted  beneath  the  raised  lower 
sash. 

Ceiling  or  wall-mounted  units  may  be  used  in  commercial  establish- 
ments, when  floor  space  is  at  a  premium.  They  generally  secure  refri- 
geration from  a  remotely  placed  condensing  unit  and  are  designed  for 
support  by  means  of  hanger  rods.  It  is  often  possible  to  conceal  them  in 
adjoining  closets  or  workrooms,  with  the  air  discharge  louvers  fixed  in  the 
intervening  wall ;  this  makes  it  easy  also  to  conceal  the  piping  connections 
and  wiring.  In  stores,  suspended  type  units  may  conveniently  be  placed 
over  the  housed-in  show  windows. 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

Portable 

Portable  summer-function  units  mounted  on  rubber-tired  casters  or 
rollers  can  be  obtained  in  the  smaller  capacities  up  to  about  9000  Btu  per 
hour.  They  may  have  built-in  compressor  units  using  city  water  for 
jacket  and  condenser  cooling,  or  they  may  employ  ice  or  low  temperature 
city  water.  Hose  connections  for  water  supply  and  return,  and  for  con- 
densate  drains,  are  needed  in  addition  to  a  plug-in  electrical  connection. 
It  is  expected  that  a  market  for  such  units  can  be  developed  in  hospitals, 
hotel  guest  rooms,  and  residences. 

Special  Types  of  Units 

The  field  for  unit  conditioners  has  been  extended  by  the  appearance  of 
small  low-cost  devices  for  comfort  cooling  that  localize  the  cooling  effect 
to  the  immediate  vicinity  of  the  user.  These  include  bed  tents  and  robe- 
type  coolers,  which  require  motors  not  larger  than  J^  hp.  The  tent  is 
suspended  from  a  bracket  attached  to  the  bed  frame,  and  the  cooler 
placed  alongside  is  connected  to  it  with  a  short  collar  for  the  air  discharge. 
The  robe-type  device  is  intended  for  barber  and  beauty  shops;  it  works 
on  the  same  principle.  Besides  handling  much  smaller  quantities  of  air, 
these  expedients  achieve  economy  because  they  operate  only  when 
required  for  the  comfort  of  the  user. 

Reversed  Refrigeration  Heating 

All-year  unit  conditioners  that  utilize  their  refrigeration  apparatus  for 
winter  heating  by  the  principle  known  as  reverse  refrigeration  cycle,  are 
being  developed.  A  detailed  explanation  of  this  system  is  given  in 
Chapter  39.  For  regions  rarely  having  winter  temperatures  below 
freezing,  there  is  believed  to  be  a  considerable  field  of  application  for  such 
equipment.  The  heat  delivered  to  the  room  will  range  between  2.5  and 
3.5  times  the  equivalent  of  the  electrical  power  taken  by  the  motor, 
depending  on  the  outdoor  temperature.  The  gain  is,  of  course,  derived 
from  the  ambient  air,  requiring  an  inlet  and  an  outlet  duct  for  passing  a 
considerable  volume.  Lower  rates  for  energy  may  sometimes  be  obtained, 
under  the  resulting  improved  annual  load  factor,  when  both  cooling  and 
heating  are  provided  electrically. 

LOCATION  OF  UNITS,  AIR  FLOW  PATHS 

The  number  of  units,  the  availability  of  space,  and  the  convenience  of 
making  piping,  wiring,  and  duct  connections,  which  involves  the  location 
of  outside  cooling,  heating,  and  power  sources,  must  be  considered  in 
choosing  locations  for  the  units,  as  must  the  positions  of  persons,  furni- 
ture, and  materials  in  the  space  to  be  conditioned,  and  the  requirements 
of  air  distribution. 

The  most  important  of  these  considerations  is  air  distribution,  and  units 
should  be  so  located  as  to  secure  uniformity  in  all  parts  of  the  room 
whether  the  application  is  for  comfort  conditioning  or  for  industrial  uses. 
The  discharge  of  cooled  air,  in  general,  should  be  upward  immediately  at 
the  conditioner,  with  sufficient  horizontal  component  to  carry  to  the  most 
remote  point;  return  to  the  inlet  of  the  unit,  which  occurs  below  the 

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CHAPTER  12 — UNIT  AIR  CONDITIONERS  AND  CONDITIONING  SYSTEMS 

breathing  line  and  along  the  floor,  should  be  at  low  velocity.  The  location 
of  doorways,  air  vents,  and  sources  of  heat  should  be  studied,  as  they  have 
a  marked  effect  on  air  flow  and  on  temperature  uniformity.  Infiltration 
through  leaky  windows  with  certain  wind  directions  likewise  disturbs  or 
restricts  the  circulation  of  air  from  the  unit  conditioner,  and  frequently 
causes  cold  spots  by  preventing  diffusion  at  the  ceiling.  Velocities  below 
the  breathing  line  should  be  kept  low — not  over  40  to  70  feet  per  minute ; 
in  this  range,  an  anemometer  will  not  work,  and  the  Kata  thermometer 
must  be  used  for  testing  purposes. 

CONSTRUCTION  OF  APPARATUS 
Description  of  Typical  Units 

The  types  and  designs  of  air  conditioning  units  proposed  or  in  produc- 
tion are  legion ;  new  designs  are  constantly  appearing,  with  a  tendency 
toward  better  mechanical  construction  and  a  wider  range  of  application. 
However,  nearly  all  types  now  commercially  available  utilize  mechanical 
refrigeration  or  cold  water  for  summer  cooling,  and  consequently  the 
descriptions  below  are  limited  to  such  equipment,  using  electric  power. 
Illustrations  of  current  makes  and  models  will  be  found  in  the  Catalog 
Data  Section  of  this  volume. 

Fig.  7  shows  an  all-year,  floor- type  unit  for  direct  expansion  of  re- 
frigerant supplied  by  a  remotely  located  compressor;  with  modifications, 
the  cooling  coil  can  be  used  with  chilled  water.  The  fans  below  the 
separate  cooling  and  heating  elements  deliver  the  air  against  deflectors 
that  give  distribution  across  the  element  face,  and  the  usual  drip  pan  for 
condensation  is  provided.  Separate  elements  for  heating  and  for  cooling 
possess  the  advantage  of  allowing  the  former  to  be  connected  to  the 
source  of  heat  with  piping  entirely  separate  from  the  refrigerant  lines  to 
the  cooling  element,  with  no  cross-connections.  Thus  the  unit  may  be 
used  for  warming  in  the  morning  and  for  cooling  later  in  the  day,  if 
desired,  without  manipulation  of  valves.  When  this  unit  is  installed  for 
cooling  only,  the  heating  element  is  omitted. 

A  summer-function  unit  with  fans  above  the  cooling  element  is  shown 
in  Fig.  6;  a  condensing  unit,  with  schematic  diagram  of  refrigerant  piping 
and  wiring,  is  included.  This  air  conditioning  unit,  as  well  as  that  in 
Fig.  7,  when  housed  in  an  ornamental  cabinet,  is  suitable  for  high  grade 
residential  or  commercial  installations. 

An  entirely  different  arrangement,  shown  in  Fig.  8,  places  both  the  air 
inlet  and  the  discharge  at  the  top  of  the  unit.  The  fan  in  the  upper 
portion  at  one  side  discharges  the  air  toward  the  bottom,  where  it  turns 
and  passes  horizontally  through  an  air  washer  equipped  with  atomizing 
sprays.  The  path  continues  vertically  upward  through  eliminators, 
cooling  surface,  and  heating  surface  before  leaving  the  unit.  With  steam 
or  hot  water  connected  to  the  heating  element,  tempered  water  to  the 
sprays,  and  refrigerated  water  to  the  cooling  element,  this  unit  gives  con- 
trolled temperature,  humidity,  air  cleaning  and  air  movement  in  both 
summer  and  winter.  Air  washing  may  be  continued  in  summer  to 
remove  room  odors'.  Acoustical  treatment  of  the  housing  and  the  outlet 
baffles  permits  installation  where  the  noise  requirements  are  exacting. 

One  of  the  most  recently  developed  units,  designed  particularly  for  low 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

cost  installations,  is  shown  in  Fig.9.  The  twin  fans  with  wheels,  mounted 
on  extensions  of  the  motor  shaft,  take  air  from  the  floor  and  send  it 
downward  through  a  passage  containing  a  water  spray.  The  direction  of 
flow  is  then  reversed,  the  air  passing  through  a  double  set  of  coils  for 
cooling  or  heating,  and  leaving  the  cabinet  through  a  top  grille.  The 


FIG.  7.  FLOOR  UNIT  FOR  HEATING  AND  COOLING 

7"o    ffo  om 

\    \        \    \    t   tt 


FIG.  8.  UNIT  WITH  TOP  INLET  AND  OUTLET 


spray  nozzles  are  supplied  with  city  water,  the  excess  collecting  in  the  air 
reversal  chamber,  which  has  a  drain.  The  cooling  coil  uses  water  from 
the  city  mains  or  other  low- temperature  source;  alternatively,  direct 
expansion  of  refrigerant  from  a  motor-driven  condensing  unit  can  be 
utilized.  The  unit  provides  all-year  functions,  the  cleaning  being  accom- 

208 


CHAPTER  12 — UNIT  AIR  CONDITIONERS  AND  CONDITIONING  SYSTEMS 

plished  by  the  spray  in  winter  and  by  contact  with  the  wetted  cooling  coil 
in  summer.  Automatic  electrically  operated  controls  for  water  flow, 
steam  flow,  temperature,  and  humidity  are  optional.  «. 

For  industrial  applications,  the  floor-type  unit,  Fig.  10,  or  the  ceiling- 
type,  Fig.  11,  may  be  used.  The  former  has  a  galvanized  steel  casing  that 
encloses  the  cooling  and  heating  elements,  with  fans  mounted  above  them ; 
the  air  discharges  from  the  top  through  90-degree  elbow  ducts,  which 
deliver  it  in  a  nearly  horizontal  direction.  The  operating  motor  for  the 
fan  is  carried  on  a  bracket  at  one  side,  and  at  the  bottom  a  condensate 
drip  pan  is  provided ;  space  between  the  pan  and  motor  bracket  is  utilized 


•Outlets 


FIG.  10.    INDUSTRIAL  FLOOR  TYPE 


FIG.  9.    ALL-YEAR  TYPE  UNIT 
CONDITIONER 

for  traps  and  valves.  This  unit  does  not  wash  or  filter  the  air,  nor  is  a 
fresh-air  supply  provided  for  ventilation;  thus  only  cooling,  dehumidi- 
fication,  and  circulation  can  be  accomplished  in  summer,  and  heating  and 
circulation  in  winter. 

The  ceiling  type,  Fig.  11,  is  primarily  for  summer  use,  although  when 
supplemented  by  a  regular  heating  system  it  can  accomplish  a  limited 
amount  of  humidifying  in  winter.  The  apparatus  consists  of  an  air  washer 
with  the  usual  water  sprays,  eliminator  plates,  and  air  circulating  fan, 
designed  for  suspension  from  the  ceiling.  The  air  supply  is  taken  from 
the  room  through  the  intake  register,  passes  through  the  water  spray  and 
eliminators,  and  is  delivered  back  into  the  room  through  the  discharge 
outlet  equipped  with  adjustable  lowers.  The  refrigeration  unit  may  be 
treated  at  any  convenient  point  and  tiie  cooled  water  circulated  to  and 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

from  the  conditioner  through  pipes  at  the  ceiling,  so  that  no  floor  space  is 
lost.  This  style  of  unit  is  for  industrial  and  large  office  installations. 
Where  the  appearance  on  the  ceiling  is  objectionable,  the  unit  may  be 
placed  at  some  other  location,  using  a  compact  duct  system  for  the  air  to 
and  from  the  conditioning  unit. 

In  the  following  paragraphs,  typical  forms  of  construction  are  outlined. 
Many  variations  of  these  have  been  used,  and  modifications  or  entirely 
new  details  are  constantly  being  introduced.  The^  Catalog  Data  Section 
illustrates  and  describes  the  current  designs  of  leading  equipment  manu- 
facturers. 

Cabinets,  Registers 

Cabinets  are  made  of  furniture  grade  sheet  steel  suitable  for  pressing  in 
panels,  protected  by  corrosion-resistant  priming  coatings.  The  design  is 
such  as  to  permit  access  to  the  equipment,  which  is  independently  sup- 
ported on  a  frame  or  chassis.  Heat  insulation  of  either  rigid  or  flexible 


Eliminators 


Intake. 


Fein 


-Pra/n 
FIG.  11.  SUMMER  COOLING  UNIT 


Chamber 


type,  to  prevent  sweating  in  summer  or  overheating  in  winter,  is  used, 
particularly  with  thermostatic  controls  that  start  and  stop  the  fans 
without  affecting  the  supply  of  heating  or  cooling  medium  to  the  coils. 
Sound-deadening  is  equally  important,  to  avoid  vibration  or  drumming 
effect  of  the  panels.  The  finish  of  commercial  and  residential  units  is 
usually  in  imitation  of  wood  grain,  or  may  be  in  solid  color  to  harmonize 
with  room  finish  and  furnishings. 

Outlet  registers  are  generally  placed  in  the  top  of  the  cabinet  to  direct 
the  air  at  an  angle  approximately  30  degrees  from  the  vertical^  They 
should  be  proportioned  to  maintain  sufficiently  high  air  velocity  for 
preventing  a  local  cold  spot  caused  by  too  short  a  flow  circuit  in  the  room. 
Types  that  give  ejector  action,  entraining  some  room  air  and  propelling 
the  mixture  a  considerable  distance  away  from  the  unit,  are  preferred. 
Return-rair  registers  should  act  as  sound-deadeners  and  serve  to  hide  the 
internal  mechanism. 

Motors 

Motors  are  usually  of  the  capacitor  or  repulsion-induction  types,  single- 
phase.  However,  in  sizes  5  hp  and  larger,  three-phase  will  ordinarily  -be 
preferable;  this  will  deperid  on  character  and  capacity  of  service  facilities 

210 


CHAPTER  12  —  UNIT  AIR  CONDITIONERS  AND  CONDITIONING  SYSTEMS 

available.  Special  designs  giving  low  starting  current,  silent  running  and 
(in  the  case  of  compressor  motors)  high  starting  torque,  are  essential. 
Features  to  minimize  lamp  flicker  and  radio  interference  must  be  incor- 
porated. coordinated  with  characteristics  of  the  compressor.  For  auto- 
matically controlled  units,  two  motors  (with  sequence  relay  for  con- 
secutive starting)  are  sometimes  direct-connected  to  the  load,  for  holding 
the  current  inrush  to  a  low  value,  when  the  starting  torque  of  the  driven 
equipment  permits.  Devices  known  as  suction  unloaders,  permitting 
large  air  or  refrigeration  compressors  to  come  up  to  speed  without  load, 
involve  too  much  complication  for  the  size  of  apparatus  used  with  unit 
conditioners. 

Refrigerants 

The  choice  of  refrigerants  for  a  direct-expansion  system  is  limited  to 
non-toxic,  nearly  odorless  fluids  —  principally  methyl  chloride,  freon  or 
iso-butane.  Local  ordinances  and  fire  regulations  prescribe  the  maximum 
quantity  of  refrigerant  in  a  system  for  residential  and  usual  commercial 
requirements.  Indirect  systems  may  use  ammonia,  sulphur  dioxide  or 
carbon  dioxide,  since  the  equipment  and  piping  can  then  be  isolated, 
remote  from  the  conditioned  rooms. 

Compressors,  Condensers,  Cooling  Coils  or  Evaporators 

Compressors  of  the  multi-cylinder  reciprocating  or  rotary  designs  are 
preferred,  as  they  minimize  starting  troubles  and  lamp  flicker.  Gland  or 
shaft-seal  leaks,  with  freon  or  methyl  chloride,  must  be  provided  against, 
because  of  the  difficulty  in  detecting  leaks  before  the  refrigerant  charge  is 
lost  ;  this  is  especially  important  when  the  pressure  in  the  crankcase  tends 
to  rise  after  the  compressor  shuts  down.  V-belt  drives  from  motors  permit 
the  compressor  and  motor  each  to  run  at  its  most  economical  speed,  and 
provide  desirable  resilience  at  the  instant  of  starting. 

Condensers  for  water  cooling  are  of  either  the  double-tube  or  shell-and- 
tube  types,  with  the  latter  preferred  when  the  water  carries  dissolved  or 
suspended  solids;  provision  for  opening  and  cleaning  should  be  made. 
Air  cooled  condensers  usually  are  supplied  with  air  by  propeller  fans 
integral  with  the  compressor  flywheels  or  mounted  on  the  compressor 
shafts. 

Evaporator  coils,  in  units  using  direct  expansion  of  the  refrigerant,  also 
constitute  the  cooling  coils  over  which  the  air  flows  to  be  cooled  and 
dehumidified.  They  are  constructed  of  metal  suitable  for  the  refrigerant 
used,  and  have  fijis-oa  the  exterior  to  increase  the  heat  transfer  per  unit 
length  of  tube.  The  arrangement  and  amount  of  surface  provided,  in 
relation  to  the  maintained  refrigerant  temperature,  the^initial  tempera- 
ture and  dew  point  of  the  air,  and  the  rate  of  air  circulation  over  the  coil 
determine  the  final  air  temperature  and  thus  the  amount  of  debumidj- 
fication  secured.  With,  indirect  refrigerating  systems,  the  cooling  fcoifo  in 
the  units  are  usually  somewhat  larger,  because  the  cooling  fluid  (water  or 
brine)  is  at  a  higher  ialet  temperature;  the  evaporator  in  this  case  is 
remotely  located  (with  tibe  <x>acbn^ia^  unit)  and  serves  to  chil  the  water 
circulated  by  a  pump  ta,  the  cofeia  £he  mnit^o^dit^i^rs> 

<m  coo&ig  mis  weqtares  &  drip  pan,  with 
f  disposal:  f>efati  tfiefc  fibteBj&by  .^jter^ryr^iriationsy  or  an 


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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

ejector  operated  from  the  city  water  supply  used  for  winter  humidi- 
fication  or  for  cooling  the  condenser.  In  some  cases,  a  condensate  storage 
tank  to  be  emptied  manually,  or  a  motor-driven  pump,  is  supplied. 
Eliminator  baffles  may  be  provided  immediately  below  the  outlet  grille  to 
intercept  any  drops  of  water  picked  up  by  the  air  current. 

Humidifiers 

Humidification  in  unit  conditioners  may  be  accomplished  by  sprays 
using  cold  or  heated  water  at  city  main  pressure,  or  by  water  trickling 
over  heated  surfaces  or  a  mesh  filling.  The  design  must  provide  for  sup- 
plying the  heat  of  evaporation,  and  for  exposing  to  the  air  current  a 
sufficient  area  of  water  film.  This  requires  a  considerable  excess  of  water, 
which  may  be  wasted  to  a  drain  or  recirculated  by  a  pump ;  with  the  latter, 
periodic  flushing  of  the  system  must  be  practicable,  to  dispose  of  the  dirt 
removed  from  the  air.  When  considerable  amounts  of  fresh  air  are 
provided  by  the  unit  conditioner  or  enter  by  infiltration,  the  quantity  of 
water  to  maintain  the  desired  humidity  is  greater  than  when  the  unit 
merely  recirculates  and  the  room  has  only  moderate  leakage.  In  the 
latter  case  the  humidifier  can  be  small,  since  only  a  slight  amount  of 
moisture  is  supplied  to  the  air  with  each  passage  through  the  unit. 

Fans,  Fresh  Air  Supply 

Fans  are  usually  of  the  centrifugal  type  with  scrolls,  inlet  cones,  blades, 
and  tip  speeds  designed  for  quiet  operation.  Compactness  and  uniform 
distribution  of  air  across  the  width  of  the  coils  and  grilles  are  obtained  by 
using  two  or  more  fans  in  parallel,  the  rotors  mounted  on  a  common  shaft 
or  on  a  double-end  extension  of  the  motor  shaft.  Housings  and  deflectors 
(if  used  at  the  fan  outlets)  may  be  acoustically  treated.  Propeller-type 
fans  are  sometimes  used,  although  more  difficult  to  make  quiet.  Efficiency 
is  a  secondary  consideration,  because  the  motors  are  of  small  fractional- 
horsepower  sizes. 

Fresh  air  supply  connections  are  usually  through  a  fixed  panel  inserted 
in  a  window  frame  between  sill  and  lower  sash;  this  has  a  louvered  and 
screened  opening,  connected  with  a  metal  duct  to  the  space  in  the  cabinet 
at  the  inlet  side  of  the  fans.  A  manually  adjustable  damper  regulates  the 
proportionate  volumes  of  recirculated  and  fresh  air. 

Filters 

Air  cleaning  devices  include  filters  of  glass  or  metal  wool,  cellulose,  felt, 
or  woven  fabric;  they  are  usually  of  the  renewable  cartridge  type,  designed 
for  low  air  resistance.  Types  especially  effective  in  the  removal  of  hay 
fever  pollen  are  desirable.  An  alternative  device  is  a  water  spray  also 
serving  as  a  humidifier  in  winter,  or  when  supplied  with  chilled  water,  as 
a  cooler  and  dehumidifier  in  summer.  The  fins  on  cooling  coils,  auto- 
matically wetted  by  condensate  obtained  in  dehumidifying  the  air,  are 
also  employed  in  some  types.  Complete  removal  of  tobacco  smoke  is  not 
possible  with  any  type  of  filter  or  washer  used  in  unit  conditioners ;  the 
limited  amount  of  ventilation  air  in  summer,  admissible  from  the  oper- 
ating cost  standpoint,  often  results  in  a  smoke  haze.  The  only  remedy  is 
increased  ventilation,  with  consequent  higher  operating  expense;  the  air 

212 


CHAPTER  12 — UNIT  AIR  CONDITIONERS  AND  CONDITIONING  SYSTEMS 

outlets  should  be  near  the  ceiling,  to  tap  the  upper  stratum  where  the 
smoke  is  most  dense. 

Heating  Coifs 

Heating  coils  are  generally  of  the  extended-surface  type  for  compact- 
ness, and  may  be  designed  for  any  pressure  of  steam  available,  or  for 
forced  circulation  hot  water — usually  about  180  F.  Some  types  of  heating 
systems,  both  hot  water  and  steam,  are  unsuited  to  unit  conditioners, 
especially  when  thermostatically  controlled,  because  of  the  resulting 
sudden  changes  in  load.  Gravity-circulation  hot  water  systems  must  be 
converted  to  forced  circulation,  if  unit  conditioners  are  to  be  connected, 
and  steam  systems  must  always  operate  at  pressures  above  atmosphere. 

Manual  and  Automatic  Controls 

Manual  control  is  generally  used  with  unit  conditioners,  because  of  the 
high  cost  of  reliable  automatic  controls.  Fan  and  compressor  motors  are 
started  and  stopped  by  individual  switches.  Fluids  for  the  heating  and 
cooling  coils  are  regulated  with  manual  valves,  generally  permitting  the 
flow  to  continue  regardless  of  whether  the  fan  is  operating;  with  this 
arrangement,  adequate  heat  insulation  must  be  provided  within  the 
cabinet,  and  the  size  of  the  unit  conditioner  in  a  room  is  limited  to  that 
which  will  give  the  minimum  required  heat  supply  by  gravity  air  circu- 
lation through  the  conditioner  when  the  fan  is  stopped. 

Automatic  controls  consist  of  a  thermostat  for  room  temperature  and  a 
hygrostat  for  humidity.  The  former  starts  and  stops  the  fan  in  the  unit 
conditioner,  thereby  controlling  the  supply  of  cooled  or  warmed  air  to  the 
room.  A  hygrostat  is  not  usually  supplied,  because  of  high  cost  and 
imperfect  reliability  of  types  now  available ;  when  used,  it  is  connected  to 
the  valve  admitting  water  to  a  spray-  or  trickle-type  humidifier,  or  to  a 
refrigerant  supply  valve  controlling  a  supplementary  section  of  the  cooling 
coil.  The  best  arrangement  is  one  that  permits  the  full  capacity  of  the 
compressor  to  be  utilized  for  either  sensible  heat  removal  or  dehumidi- 
fication,  based  on  the  principle  that  the  compressor  capacity  varies 
approximately  as  the  temperature  of  the  refrigerant  in  the  cooling  coil. 
Compressors  are  started  and  stopped  by  pressure  switches  on  the  dis- 
charge (high  pressure)  side.  Water  supply  to  the  compressor  jackets  and 
the  compressor  is  turned  on  and  off  by  a  solenoid  valve  energized  when 
the  compressor  motor  starts.  Refrigerant  supply  to  the  cooling  coil 
(constituting  the  evaporator)  is  usually  regulated  by  a  thermostatic  valve, 
as  a  function  of  the  refrigerant  outlet  temperature,  or  by  a  flow  valve  that 
tends  to  hold  a  constant  level  in  the  liquid  receiver. 

INSTALLATION  OF  UNIT  CONDITIONERS 
Piping,  Wiring,  Ducts 

Piping  connections  for  water  and  steam  are  made  preferably  with 
corrosion  resisting  material,  usually  brass  or  copper.  Light  weight  rigid 
tubing  with  sweated  joint  fittings  has  advantages  over  threaded  construc- 
tion. Flexible  copper  tubing  with  compression  type  connections  may  be 
used  in  the  smaller  sizes  (u>f>  to  %  in.  dra.),  as  it  lends  itself  to  conceal- 
ment In  existing  walls  of  other  places  difficult  of  access ;  distribution  of  the 

213 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

refrigerant  from  a  remotely  located  condensing  unit  is  usually  made  with 
such  flexible  tubing. 

Wiring  connections  should  be  made  using  modern  materials  and 
methods,  such  as  will  receive  approval  of  local  inspection  authorities 
having  jurisdiction.  For  portable  conditioning  units  with  built-in  com- 
pressors, particular  care  should  be  taken  to  select  rugged  receptacles  and 
plugs;  waterproof  flexible  cords  are  recommended  because  of  the  pos- 
sibility of  water  leakage  from  adjacent  hose  connections  or  by  overflow 
from  the  unit  if  the  drain  becomes  stopped. 

Ducts  for  outgoing  air  supply,  usually  from  nearby  window  openings, 
present  no  particular  problems. 

Workmanship 

The  requirements  as  to  workmanship  for  installation  of  unit  con- 
ditioners are  exceptionally  severe;  this  is  particularly  true  for  work  in 
high  grade  offices  and  residences,  in  occupied  quarters.  Handling  of  the 
materials  and  the  cutting,  patching,  and  refinishing  invariably  demand 
neatness,  accuracy,  and  planning  that  the  ordinary  mechanic  is  un- 
familiar with,  so  that  close  supervision  must  be  given. 

BASIS  OF  EQUIPMENT  RATINGS1 

While  no  uniform  standard  for  rating  unit  air  conditioners  has  yet  been 
adopted,  manufacturers  generally  give  a  definite  rating  for  each  size  unit, 
based  on  the  volume  of  air  handled  by  the  fan  for  cooling;  the  rating  is 
stated  in  Btu  per  hour  at  a  given  dry-  and  wet-bulb  temperature  of  air 
entering  the  unit,  with  a  given  refrigerant  temperature  maintained  within 
the  coil,  resulting  in  a  stated  relationship  between  sensible  and  latent  heat 
removal.  The  temperature  of  the  cooling  water  or  air  supply  for  the 
condensing  unit  is  also  involved.  The  duty  for  heating  service  is  likewise 
given  in  Btu  per  hour  with  70  F  room  temperature,  for  a  stated  steam 
pressure  or  hot  water  temperature  (usually  180  F) .  Humidifying  capacity 
is  based  on  hourly  weight  of  water  evaporated.  The  Catalog  Data  Section 
in  this  volume  gives  the  ratings  of  current  models  offered  by  leading 
manufacturers. 

METHODS  OF  CALCULATING  REQUIRED  CAPACITY 

In  estimating  the  load  for  unit  air  conditioning  apparatus,  a  survey 
should  be  made  of  the  surrounding  conditions  and  the  heat  quantities 
calculated.  The  climatic  conditions  representing  the  maximum  loads  to 
be  designed  for  should  be  carefully  determined. 

Cooling  Loads 

For  cooling  loads  served  by  unit  conditioners,  the  factors  for  heat  gains 
and  losses  are  the  same  as  apply  to  central  fan  systems.  The  sensible 
heat  gains  are  from  the  following  sources: 


iRefer  to  the  standard  ratings  of  air  conditioning  equipment  of  the  National  Electric  Manufactures 
Association. 

214 


CHAPTER  12 — UNIT  AIR  CONDITIONERS  AND  CONDITIONING  SYSTEMS 

a.  Sun  effect. 

b.  Transmission  through  walls,  floors,  ceilings,  glass,  and  roofs. 

c.  Infiltration,  including  ventilation  air. 

d.  People. 

e.  Lights. 

/.   Electric  motors  and  appliances. 
g.  Steam  and  gas  appliances. 
h.  Miscellaneous  heat  sources. 

The  latent  heat  load,  usually  determined  separately,  comes  from 
dehumidification  of  the  air  and  from  people  and  materials.  The  method 
and  factors  to  be  used  are  outlined  in  standard  texts  and  in  manufacturers* 
handbooks. 

Rated  capacity  for  direct-expansion  refrigeration  units  should  include 
an  allowance  for  the  heat  equivalent  of  the  fan-motor  input,  plus  the 
portion  of  the  power  to  the  compressor  not  removed  by  the  cooling  water. 
For  indirect-expansion  systems,  allowance  should  be  made  for  heat 
pickup  by  the  refrigerant  circulating  lines,  or  for  the  pickup  by  a  chilled- 
water  or  brine-storage  tank  and  for  the  shaft-horsepower  input  to  a 
circulating  pump,  if  used. 

As  a  rough  approximation,  the  refrigeration  tonnage  required  for  unit 
conditioners  serving  rooms  devoted  to  various  uses  may  be  assumed  as 
follows : 


TYPES  OF  ROOMS 

CTT  FT  PER  TON 

Cafeterias,  lunchrooms  

1000  to  1500 

Barber  and  beauty  shops,  dance  halls 

1200  to  1800 

Dining  rooms,  crowded  retail  stores  

1500  to  2000 

Theaters  

1800  to  2400 

General  offices,  club  rooms,  retail  stores,  funeral  parlors     

2000  to  3000 

Banks,  brokers'  offices,  private  offices,  residences 

2500  to  4000 

Obviously,  there  will  be  many  cases  to  which  the  mentioned  limiting 
values  do  not  apply,  A  calculation  of  the  cooling  load,  based  on  an 
accurate  survey,  should  always  be  made  before  recommending  the  size  of 
an  installation  or  naming  a  cost  figure. 

Heating  Loads 

Heating  loads  are  calculated  in  the  usual  manner,  as  outlined  in 
Chapter  7.  Allowance  must  be  made  also  for  the  latent  heat  supplied  to 
the  water  for  humidification,  when  the  infiltration  or  ventilation  air 
quantity  is  large. 

APPROXIMATE  COSTS 
Equipment  and  Installation 

Floor  type  all-year,  unit  conditioners,  oon^pletely  self-e6ntaiaed  and 
equipped  with  motor-<Mve0  ^compressors  and  iiN&rm^t^tie  controls, 

'  at  the 

,  in- 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

eluding  expense  for  piping,  wiring  and  fresh-air  ducts,  amounting  to 
between  175  and  $150,  with  perhaps  $50  additional  if  overtime  work  is 
necessary  to  avoid  inconveniencing  the  occupants  of  offices  or  other 
quarters. 

A  similar  unit  without  compressor,  using  chilled  water  or  direct- 
expansion  refrigerant  from  a  remotely  located  compressor,  costs  $175  or 
more  at  the  factory.  Installation  expense  is  somewhat  greater  than  for 
the  self-contained  unit,  because  the  refrigerant  piping  costs  more  than  is 
saved  by  the  reduction  in  wiring.  Omission  of  the  heating  coil,  confining 
the  unit  to  summer  functions  only,  lowers  the  price  by  $25  to  $100. 

For  smaller  units,  rated  between  6000  and  8000  Btu  per  hour,*providing 
all-year  service  and  equipped  with  motor-driven  compressors,  the  price 
ranges  from  $325  to  $450  at  the  factory.  Larger  units,  rated  at  about 
24,000  Btu  per  hour  for  cooling,  cost  between  25  and  45  per  cent  more 
than  the  12,000-Btu  per  hour  size.  Delivery  and  installation  expense 
for  either  of  these  sizes  does  not  differ  more  than  25  per  cent  from  that 
of  the  12,000-Btu  per  hour  unit. 

Industrial- type  conditioners,  either  floor  or  suspended  models,  are 
usually  made  only  in  ratings  of  20,000  Btu  per  hour  and  higher;  omission 
of  expensively  finished  cabinets  and  other  differences  reduces  the  cost  con- 
siderably below  that  of  corresponding  sizes  of  commercial  and  residential 
types. 

Condensing  units  completely  assembled  on  bedplates,  especially 
adapted  to  serve  one  or  more  unit  conditioners,  are  available.  They  com- 
prise a  motor,  compressor,  condenser,  liquid  receiver,  and  control  devices, 
and  they  are  arranged  for  water  cooling  or  are  equipped  (in  the  smaller 
sizes)  with  fans  for  air  cooling.  Prices  for  representative  sizes,  including 
motors  but  not  starting  equipment,  are  as  follows: 


BTU  PER  HOUR 

FACTORY  PRICE 

INSTALLATION  COST 

8,000 
12,000 
36,000 
60,000 
120,000 

$275  and  up 
325  and  up 
575  and  up 
800  and  up 
1100  and  up 

$60  and  up 
65  and  up 
80  and  up 
90  and  up 
125  and  lip 

These  prices  are  for  water-cooled  types;  air  cooling  adds  $25  to 
Installation  cost  includes  transportation,  foundations,  wiring,  starting 
equipment,  cooling  water  piping  or  air  ducts,  and  sound-deadening 
insulation.  Refrigerant  connections  from  liquid  receiver  and  compressor 
suction  to  unit  conditioners  are  not  included.  For  office  buildings  and 
similar  occupied  quarters,  overtime  labor  may  increase  the  cost. 

The  prices  given  represent  net  cost  to  the  ultimate  purchaser.  Although 
roughly  indicative  of  the  present-day  market,  they  should  not  be  used  as 
a  basis  of  a  specific  estimate  or  an  appropriation,  because  designs,  ratings, 
and  prices  vary  considerably  between  makers  and  in  different  parts  of  the 
country.  Transportation  and  installation  expense  is  even  more  variable, 
depending  upon  freight  rates,  wage  scales,  and  particularly  on  the  con- 
dition of  the  building  and  the  adequacy  of  existing  piping  and  wiring 
systems  to  which  the  unit  conditioners  are  to  be  connected.  Furthermore, 

216 


CHAPTER  12 — UNIT  AIR  CONDITIONERS  AND  CONDITIONING  SYSTEMS 

the  industry  is  in  a  state  of  fairly  rapid  development,  so  that  any  general 
cost  figures  should  be  used  with  caution. 

Operation 

For  a  24,000-Btu  per  hour  unit  operating  at  full  load  for  summer 
cooling,  with  electricity  at  §0.05  per  kwh  and  70  F  city  water  at  $1.50 
per  1000  cu  ft,  the  hourly  electric  and  water  expense  works  out  to  $0.14. 
Under  climatic  conditions  representative  of  a  large  part  of  the  country, 
the  load  factor,  during  the  10  hours'  daily  operation  required,  averages 
50  per  cent;  this  gives  a  daily  operating  cost  of  $0.70.  The  seasonal  cost 
for  localities  requiring,  for  example,  1000  hours  of  operation  (at  50  per 
cent  load  factor)  then  becomes  $70.  To  this  should  be  added  main- 
tenance and  fixed  charges  of  25  per  cent  (based  on  about  a  five-year  useful 
life)  on  an  investment  around  $1200.  The  over-all  expense  for  owning  and 
operating  is  thus  of  the  order  of  $370  per  year.  Such  a  cost  may  be 
incurred,  in  a  climate  like  that  of  New  York  City,  by  the  owner  of  a  home 
in  which  at  least  the  living  room,  the  dining  room,  and  a  bedroom  are 
cooled  with  unit  conditioners  served  by  refrigerating  equipment  of  the 
mentioned  capacity. 

This  expense  may  be  compared  with  the  cost  of  winter  heating,  com- 
puted by  adding  annual  fixed  and  maintenance  charges  to  cost  of  fuel, 
attendance,  and  other  items.  The  comfort  attainable  in  hot,  humid 
weather  is  so  welcome  that  these  costs  will  undoubtedly  be  looked  upon  as 
reasonable  by  an  increasing  number  of  people,  as  they  become  personally 
familiar  with  the  value  of  the  service  rendered  by  modern  air  conditioning 
equipment.  Exposure  to  such  comfort  in  commercial  establishments, 
railroad  trains,  and  other  public  places  will  unquestionably  tend  to 
increase  the  demand  for  home  installations  at  a  greater  rate  each  year. 
The  developments  in  equipment  for  the  type  of  service  described  have 
been  rapid  during  the  past  few  years  and  the  latest  models  may  be  seen 
in  the  Catalog  Data  Section. 


PROBLEMS  IN  PRACTICE 


1  •  Are  unit  conditioners  necessarily  self-contained? 

No.  The  heating  medium  is  always  supplied  from  a  separate  plant,  and  the  refrigerant 
for  cooling  and  dehumidification  may  come  from  a  separately  located  compressor  or 
other  supply  source. 

2  0  Are  ducts  used  with  unit  conditioners? 

Yes.  Usually  a  short  connection  for  fresh  air  intake  is  made  to  an  adjacent  window 
or  wall  opening.  Occasionally  ducts  are  required  for  return  air  and  for  discharge,  when  a 
unit  is  located  near  the  room  served  but  not  within  it. 

3  •  What  is  the  meaning  of  the  term  condensing  unit  in  relation  to  unit  air 
conditioners? 

A  condensing  unit  is  the  assembly,  on  a  bedplate,  of  a  compactly  arranged  refrigeration 
compressor,  motor,  drive,  condenser,  liquid  receiver,  and  automatic  controls  used  for 
supplying  the  refrigeration. 

217 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

4  •  Why  are  metal  surface  cooling  elements  instead  of  liquid  spray  chambers 
used  in  the  design  of  most  unit  air  conditioners  and  unit  coolers? 

The  first  cost  of  the  surface  cooling  type  of  unit  is  considerably  less  than  the  cost  of 
spray  type  equipment.  Further,  the  requirements  of  many  industrial  air  conditioning 
jobs  and  of  all  comfort  cooling  jobs  where  unit  equipment  is  applicable  can  often  be 
effectively  met  with  the  use  of  surface  type  units,  with  a  reduction  in  the  jspace  required 
for  making  the  installation.  Where  space  conditions  are  especially  limited,  the  cross- 
sectional  area  of  the  surface  cooler  can  be  reduced  because  the  resulting  increase  in 
velocity  over  the  coil  surface  increases  the  effectiveness  of  the  ^  surface,  whereas  an 
increase  in  velocity  through  a  liquid  spray  would  reduce  its  effectiveness. 

5  •  Why  are  air  conditioning  units  with  metal  cooling  surfaces  not  desirable 
for  all  industrial  jobs? 

Wherever  unusually  close  control  of  relative  humidity  is  required,  a  spray  type  unit  will 
prove  to  be  more  "satisfactory.  Relative  humidity  control  and  accurate  temperature 
control,  however,  can  be  maintained  without  difficulty  with  the  use  of  metal  surface 
units. 

6  •  Why  is  accurate  control  of  relative  humidity  with  surface  coolers  more  or 
less  complicated? 

A  surface  cooler  cannot  add  moisture  to  the  air,  and  moisture  is  removed  only  when  the 
surface  temperature  is  below  the  entering  dew-point  temperature.  Any  change  in 
condition  of  the  entering  air  will  result  in  a  change  in  the  dry-bulb  depression  of  the 
leaving  air.  This  change  in  entering  condition  requires  not  only  a  readjustment  of  the 
air  volume  but  also  a  change  in  the  coil  temperature,  if  accurate  control  over  the  relative 
humidity  is  to  be  maintained. 

7  •  What  in   general   are   the   characteristics   of  unit    conditioner   operation 
using  surface  coils? 

For  a  constant  entering  dry-bulb  temperature  and  a  constant  refrigerant  temperature 
any  increase  in  the  entering  wet-bulb  temperature  will  produce  a  rise  in  the  leaving  dry- 
bulb  temperature  with  an  accompanying  reduction  in  the  wet-bulb  depression  of  the 
leaving  air.  The  sensible  heat  removed  by  the  unit  decreases  and  the  latent  heat  in- 
creases, while  the  total  heat  removed  also  increases.  When  the  dry-bulb  temperature  of 
entering  air  is  increased,  with  constant  refrigerant  temperature  and  constant  wet-bulb 
temperature  of  entering  air,  the  wet-bulb  depression  of  the  leaving  air  increases,  and 
since  it  is  this  depression  which  determines  the  maintained  relative  humidity  it  must  be 
carefully  considered  when  selecting  the  unit. 

8  •  If  a  drop  in  the  dry -bulb  temperature  of  entering  air  reduces  the  capacity 
of  the  unit,  is  there  not  danger  of  selecting  a  unit  which  is  too  small,  if  its 
selection  should  be  based  on  an  excessive  entering  dry-bulb  temperature? 

Yes.  If  the  total  cooling  load  is  largely  internal  (such  as  from  occupants  and  lights)  as 
distinguished  from  the  cooling  load  of  outdoor  air,  and  the  unit  is  selected  on  the  basis 
of  a  too  high  dry-bulb  temperature  of  entering  air,  then,  in  the  event  of  under  capacity, 
it  might  be  possible  to  maintain  the  room  temperature  by  reducing  the  quantity  of  out- 
door air.  But  this  increases  the  recirculated  air  taken  into  the  unit,  reducing  the  dry- 
bulb  temperature  of  entering  air  and,  therefore,  reducing  the  sensible  heat  capacity  of 
the  unit.  This  reduction  in  capacity  may  offset  the  gain  obtained  by  reducing  the 
amount  of  outdoor  air  taken  in.  Further,  since  the  total  tonnage  required  for  any  instal- 
lation is  equal  to  the  total  internal  heat  load  plus  the  total  heat  removed  from  the  out- 
door air,  and  since  the  outdoor  air  might  have  a  wet-bulb  temperature  equal  to  the 
designed  wet-bulb  but  less  than  the  designed  dry-bulb  temperature,  then  the  sensible 
heat  capacity  of  the  unit  will  be  less  than  that  required.  It  follows  that  unit  air  con- 
ditioners and  coolers  should  not  be  selected  on  a  basis  of  the  maximum  possible  dry-bulb 
temperature  of  entering  air. 


218 


Chapter  13 

UNIT  HE ATKKS.  VENTILATORS, 
AND  COOLERS 

Types  of  Unit  Heaters,  Heating  Media,  Entering  and  Delivery- 
Temperature,  Output  of  Unit  Heaters,  Direction  of  Discharge, 
Boiler  Capacity,  Direct-Fired  Units,  Unit  Ventilators,  Split  and 
Combined  Systems,  Location  of  Unit  Ventilators,  Capacities, 
Attic  Fans,  Unit  Coolers 

A  UN  IT  heater  consists  of  the  combination  of  a  heating  element  and  a 
fan  or  blower  having  a  common  enclosure,  and  placed  within  or 
adjacent  to  the  space  to  be  heated.  Generally,  no  ducts  are  attached  to 
the  inlets  or  outlets.  A  unit  ventilator  is  similar  in  principle  of  operation 
to  a  unit  heater,  but  is  designed  to  use  all  or  part  outdoor  air  with  or 
without  alternate  provision  for  handling  recirculated  air.  Unit  heaters 
are  designed  mainly  for  factory  and  industrial  use,  whereas  unit  venti- 
lators are  intended  largely  for  school  and  office  ventilation  and  heating. 

Unit  heaters  and  unit  ventilators  are  designed  to : 

1 .  Circulate  the  air  in  the  building  at  a  rapid  rate. 

2.  Reduce  the  temperature  differential  between  floor  and  ceiling. 

3.  Direct  the  heated  air  so  as  to  accomplish  the  positive  and  rapid  placing  of  the 
heat  where  it  is  effective. 

4.  Remove  the  cold  stratum  of  air  from  the  floor. 

TYPES  OF  UNITS 

There  are  many  types  of  unit  heaters  available.  Most  of  them  employ 
convectors  to  be  supplied  with  steam  or  hot  water.  Some  are  mounted  on 
the  floor,  whereas  others  are  designed  for  suspension  overhead.  Heating 
surfaces  in  the  form  of  steel  pipe  coils,  non-ferrous  tubes  or  shapes  with 
extended  surfaces,  cast-iron,  and  pressed  and  built-up  sections  of  the 
cartridge  or  automotive  type  are  all  used  in  unit  heater  construction. 

Among  the  unit  heaters  available  are  types  designed  especially .  for 
industrial  purposes  having  from  one  to  four  warm  air  outlets  per  heater 
which  may  be  arranged  to  discharge  in  selected  directions  and  which  will 
project  their  heating  effects  over  distances  of  from  30  to  200  ft  from  the 
heater,  depending  upon  the  capacity  of  the  heater  and  upon  the  design  of 
the  fan  and  outlets.  Because  these  heaters  have  been  satisfactory  when 
placed  as  far  as  400  ft  from  each  other,  it  is  possible  to  select  the  heater 
location  best  suited  to  the  production  layout  in  factories.  There  are 
available  propeller  fan  type  heaters  of  smaller  capacity  with  outlet 
velocities  of  from  300  to  800  fpm,  and  these  may  be  placed  from  30  to 
100  ft  apart. 

219 


AMERICAN  SOCIETY  o/  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


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220 


CHAPTER  13 — UNIT  HEATERS,  VENTILATORS,  AND  COOLERS 

HEATING  MEDIA 

The  convectors  in  unit  heaters  or  ventilators  may  be  supplied  with 
either  hot  water  or  steam.  When  water  is  used,  it  should  be  circulated 
mechanically,  and  the  pumpage  rate  and  friction  loss  should  be  based 
upon  test  data  from  the  particular  unit  to  be  employed.  The  heat  output 
of  a  given  heater  will  be  less  when  using  water  than  with  steam,  even  at 
the  same  temperature. 

Either  high  or  low  pressure  steam  may  be  used,  but  the  proper  venting 
of  air  and  the  prevention  of  flash  steam  from  the  condensate  in  the 
returns  become  increasingly  troublesome  as  the  steam  pressure  increases. 
The  use  of  properly  constructed  traps  with  some  reliable  form  of  ther- 
mostatic  air  by-pass  solves  the  first  of  these  problems,  while  proper 
venting  or  the  use  of  condensing  legs  solves  the  second.  Increasing  the 
return  temperature  tends  to  increase  return  line  corrosion,  especially  at 
points  where  overheated  condensate  or  steam  are  led  into  a  line. 

When  low  pressure  steam  is  used  with  unit  heaters  and  ventilating 
units  it  is  highly  important  that  proper  means  be  provided  for  taking  care 
of  the  heavy  condensation.  They  should  not  be  applied  to  low  pressure 
gravity  return  systems  except  where  the  difference  between  the  heater 
level  and  the  boiler  water  line  is  large  enough  to  compensate  for  the 
pressure  loss  through  the  convector  at  its  highest  possible  condensation 
rate.  The  use  of  vacuum  or  return  pumps  and  receivers  is  advisable, 
with  jobs  of  any  considerable  size,  as  the  surest  way  of  taking  care  of 
condensate  and  at  the  same  time  providing  for  proper  venting  of  the 
units  directly  into  a  vacuum  return  line  system,  or  into  an  open  vented 
return  system,  the  latter  having  some  advantage  in  preventing  the 
formation  of  any  vacuum  in  the  unit  itself,  which  sometimes  tends  to  hold 
up  condensate  and  cause  freezing. 

ESTIMATING  HEAT  LOSSES 

The  heat  losses  of  a  building  to  be  equipped  with  unit  heaters  are 
determined  in  the  same  manner  as  for  any  other  heating  system,  excepting 
so  far  as  the  unit  heaters  may  prevent  air  stratification  and  thus  reduce 
the  temperature  difference  between  the  ceiling  and  floor.  (See  Chapter  7.) 

Unit  heaters  may  be  arranged  to  recirculate  the  air  or  to  supply  warmed 
air  from  the  outside  for  ventilation  or  to  make  up  air  exhausted. 

If  all  or  a  part  of  the  air  is  to  be  taken  in  from  out-of-doors,  the  heat 
necessary  to  warm  this  air  from  the  outside  temperature  to  the  inside 
temperature  must  be  added  to  the  transmission  or  other  losses.  Units  of 
the  number  and  size  needed  to  furnish  the  total  heat  required  are  then 
selected  from  the  manufacturers'  rating  tables,  using  these  ratings  at  the 
steam  pressure  to  be  used  and  at  the  temperature  at  which  the  air  will 
enter  the  convector. 

AIR  TEMPERATURES 

For  recirculating  heaters  with  intakes  at  the  floor  level,  the  temperature 
to  be  maintained  in  the  room  should  be  used  as  the  temperature  of  the  air 
entering  the  heater.  Where  suspended  heaters  are  used  without  any 
intake  boxes  extending  down  to  the  floor  level,  a  higher  entering  air 

221 


AMERICAN  SOCIETY  of  HEATING 

and 

VENTILATING  ENGINEERS 

GUIDE, 

1935 

TABLE  2.  CONSTANTS  FOR  DETERMINING  THE  CAPACITY  OF  Draw-THROUGH  TYPE  UNIT  HEATERS  FOR  VARIOUS  STEAM  PRESSURES 
AND  TEMPERATURES  OF  ENTERING  AIR 

(Based  on  Steam  Pressure  of  2-lb  Gage  and  Entering  Air  Temperature  of  60  F) 

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222 


CHAPTER  13 — UNIT  HEATERS,  VENTILATORS,  AND  COOLERS 

temperature  should  be  used  than  that  at  which  the  room  is  to  be  main- 
tained. With  suspended  heaters  taking  in  air  at  some  distance  above  the 
floor,  the  temperature  variation  from  floor  to  ceiling  may  reach  as  much 
as  1  deg  for  each  foot  of  elevation  during  periods  when  the  maximum 
capacity  of  the  heaters  is  required.  Unit  heaters  taking  in  recirculated 
air  at  the  floor  level  should  maintain  -temperature  differentials  of  less 
than  0.5  deg  per  foot  of  elevation  when  the  maximum  capacity  of  the 
heaters  is  required.  These  temperature  differences  per  foot  of  elevation 
are  less  than  the  corresponding  variations  per  foot  of  elevation  for  spaces 
heated  by  direct  radiation. 

Unit  heaters  save  fuel  because  of  their  ability  to  circulate  air  at  a  lower 
average  temperature  than  the  air  circulated  by  direct  radiators ;  however, 
the  unit  heaters  must  circulate  more  air  in  any  given  time  than  is  needed 
with  direct  radiators.  This  requires  the  selection  of  heaters  having  a 
liberal  air  capacity  for  the  required  heat  output,  which  in  turn  means  a 
relatively  low  final  temperature.  Extremely  low  final  temperatures  can 
be  had  only  at  the  expense  of  larger  heaters  and  increased  power,  so  that 
an  economic  limit  is  imposed.  In  general,  for  heating  purposes  it  is 
advisable  to  use  a  delivery  temperature  not  more  than  70  F  above  the 
average  room  temperature  desired,  and  one  considerably  less  where 
possible. 

OUTPUT  OF  HEATERS 

It  is  standard  practice  to  rate  unit  heaters  in  Btu  per  hour  at  a  given 
temperature  of  air  entering  the  heater  and  at  a  given  steam  pressure 
maintained  in  the  coil.  Steam  at  2  Ib  pressure  and  air  entering  at  60  F 
are  used  as  the  standard  basis  of  rating1.  The  capacity  of  a  heater 
increases  as  the  steam  pressure  increases,  and  decreases  as  the  entering 
air  temperature  increases.  The  heat  capacity  for  any  condition  of  steam 
pressure  and  entering  air  temperature  may  be  calculated  approximately 
from  any  given  rating  by  the  use  of  factors  in  Tables  1  and  2.  Table  1 
is  for  blow-through  and  Table  2  is  for  draw-through  unit  heaters.  These 
tables  are  accurate  within  5  per  cent. 

The  ratings  customarily  published  for  unit  heaters  apply  only  for 
recirculation  and  free  discharge,  unless  otherwise  noted  in  the  rating 
tables.  If  outside  air  intakes,  filters,  or  ducts  on  the  discharge  side  are 
used  with  the  heater,  proper  consideration  should  be  given  to  the  reduc- 
tion in  air  and  heat  capacity  that  will  result  because  of  this  added 
resistance. 

The  percentage  of  this  reduction  in  capacity  will  depend  upon  the 
characteristics  of  the  heater  and  on  the  type,  design,  and  speed  of  the 
fans  employed,  so  that  no  specific  percentage  of  reduction  can  be  assigned 
for  all  heaters  for  a  given  added  resistance-  In  general,  however,  disc 
or  propeller  fan  units  will  have  a  larger  reduction  in  capacity  than  housed 
fan  units  for  a  given  added  resistance,  and  a  given  heater  will  have  a 
larger  reduction  in  capacity  as  the  fan  speed  is  lowered.  When  confronted 
with  this  problem  the  ratings  under  the  conditions  expected  should  be 
secured  from  the  manufacturer. 


iSee  A.S.H.V.E.  Standard  Code  for  Testing  and  Rating  Steam   Unit  Heaters  (A.S.H.V.E.  TRANS- 
ACTIONS, Vol.  36,  1930). 

223 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

When  steam  supplied  to  the  heaters  contains  superheat,  the  capacity 
of  the  heater  will  be  but  slightly  less  than  with  saturated  steam  at  the 
same  pressure.  Recent  tests  indicate  that  the  reduction  of  capacity 
from  this  cause  is  negligible  for  superheat  up  to  50  deg  and  will  not 
exceed  3J^  per  cent  for  any  degree  of  superheat. 

Heaters  may  be  distributed  through  the  central  portions  of  a  room 
discharging  toward  exposed  surfaces,  or  may  be  spaced  around  the  walls, 
discharging  along  the  walls  and  inward  as  well,  especially  when  there  are 
considerable  roof  losses. 

In  general,  it  is  better  to  direct  the  discharge  from  the  unit  heaters 
in  such  fashion  that  rotational  circulation  of  the  entire  room  content  is 
set  up  by  the  system  rather  than  to  have  the  heaters  discharge  at  random 
and  in  counter  directions. 

DIRECTION  OF  DISCHARGE 

Various  types  and  makes  of  unit  heaters  are  illustrated  in  the  Catalog 
Section  of  this  edition.  Usually  hot  blasts  of  air  in  working  zones  are 
objectionable,  so  heaters  mounted  on  the  floor  should  have  their  discharge 
outlets  above  the  head  line  and  suspended  heaters  should  be  placed  in 
such  manner  and  turned  in  such  direction  that  the  heated  air  stream  will 
not  be  objectionable  in  the  working  zone.  In  the  interest  of  economy, 
however,  the  elevation  of  the  heater  outlet  and  the  direction  of  discharge 
should  be  so  arranged  that  the  heated  air  shall  be  brought  as  close  to 
the  head  line  as  possible,  yet  not  into  the  working  zone.  In  general,  the 
higher  the  elevation  of  the  unit,  the  greater  the  volume  and  velocity 
required  to  bring  the  warm  air  down  to  the  working  zone,  and  conse- 
quently, the  lower  the  required  temperature  of  the  air  leaving  the  unit. 

BOILER  CAPACITY 

The  capacity  of  the  boiler  should  be  based  on  the  rated  capacity  of  the 
heaters  at  the  lowest  entering  air  temperature  that  will  occur,  plus  an 
allowance  for  line  losses.  Ordinarily  for  recirculating  heaters  the  lowest 
entering  temperature  will  occur  at  the  beginning  of  the  heating  period 
and  is  usually  taken  as  40  F,  while  for  ventilators  taking  air  from  outdoors 
the  lowest  entering  temperature  will  be  the  extreme  outdoor  temperature 
expected  in  the  district.  No  greater  allowance  in  boiler  capacity  beyond 
the  calculated  heat  demand  need  be  added  in  order  to  supply  unit  heaters 
than  for  any  other  type  of  system. 

It  is  unwise  to  install  a  single  unit  heater  as  the  sole  load  on  any 
boiler,  particularly  if  the  unit  heater  motor  is  started  and  stopped  by 
thermostatic  control.  The  wide  and  sudden  fluctuations  of  load  that 
occur  under  such  conditions  would  require  closer  attendance  to  the  boiler 
than  is  usually  possible  in  a  small  installation.  Where  oil  or  gas  is  used 
to  fire  the  boiler,  it  is  possible  by  means  of  a-  pressurestat  to  control  the 
boiler,  in  response  to  this  rapid  fluctuation.  In  most  cases,  however,  and 
particularly  where  the  boiler  is  coal-fired,  it  is  advisable  to  use  two  or 
more  smaller  heating  units  instead  of  one  large  unit. 

Steam  pressures  below  5  Ib  can  be  used  with  safety  for  recirculating 
unit  heaters  when  their  coils  are  designed  for  the  purpose  and  when 

224 


CHAPTER  13 — UNIT  HEATERS,  VENTILATORS,  AND  COOLERS 

proper  provision  is  made  for  returning  the  condensate.  If  ventilators  are 
to  take  in  air  that  may  be  at  a  temperature  below  freezing,  however,  a 
steam  pressure  of  not  less  than  5  Ib  should  be  maintained  on  the  convector 
or  a  corresponding  differential  in  pressure  between  the  supply  and  returns 
be  maintained  by  means  of  a  vacuum. 


QUIETNESS 

In  selecting  unit  heaters,  attention  should  be  given  to  the  degree  of 
quietness  required  for  the  installation. 

No  given  fan  speed  may  be  applied  as  a  measure  of  relative  quietness 
to  fans  of  different  designs  and  proportions.  Quietness  is  a  function  of 
type,  diameter,  blade  form  and  other  variables  besides  speed,  and  all 


•VACUUM  BREAKER 


FLOAT  OR  BLA3T 
TRAP. 


FIG.  1.     UNIT  HEATER  CONNECTIONS 

WHERE  CONDENSATION  Is  RETURNED 

TO  VACUUM    PUMP  OR  TO  AN  OPEN 

VENTED  RECEIVER 


SUPPLY— 


VALVES  r 


•AIR    VENT  VALVE 


-WET  RETURN 


FIG.  2.  UNIT  HEATER  CONNECTIONS 
WHERE  CONDENSATION  Is  RETURNED 
TO  BOILER  THROUGH  WET  RETURN 


these  must  be  taken  into  account.  In  general  small  fans  may  be  run  at 
higher  motor  speeds  than  large  fans  with  equal  quietness ;  and  centrifugal 
fans  are  more  easily  made  quiet  than  disc  or  propeller  fans. 

PIPING  CONNECTIONS 

Piping  connections  for  unit  heaters  are  similar  to  those  for  other  types 
of  fan-blast  heaters.  Typical  connections  are  shown  in  Figs.  1  and  2. 

One-pipe  gravity  and  vapor  systems  are  not  recommended  for  unit 
heater  work. 

For  two-pipe  closed  gravity  return  systems  the  return  from  each  unit 
should  be  fitted  with  a  heavy-duty  or  blast  trap,  and  an  automatic  air 
valve  should  be  connected  into  the  return  header  of  each  unit.  Pressure- 
drop  must  be  compensated  for  by  elevation  of  the  heater  above  the  water 
line  of  the  boiler  or  of  the  receiver. 

In  pump  and  receiver  systems  the  air  may  be  eliminated  by  individual 
air  valves  on  the  heaters,  or  it  may  be  carried  into  the  returns  the  same  as 
for  vacuum  systems  and  the  entire  return  system  be  free-vented  to  the 

225 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

atmosphere,  provided  all  units,  drip  points,  and  radiation  are  properly 
trapped  to  prevent  steam  entering  the  returns. 

On  vacuum  or  open  vented  systems  the  return  from  each  unit  should  be 
fitted  with  a  large  capacity  trap  to  discharge  the  water  of  condensation 
and  with  a  thermostatic  air  valve  for  eliminating  the  air,  or  with  a  heavy- 
duty  trap  for  handling  both  the  condensation  and  the  air,  provided  the 
air  finally  can  be  eliminated  at  some  other  point  in  the  return  system. 

For  high  pressure  systems  the  same  kind  of  traps  may  be  used  as  with 
vacuum  systems,  except  that  they  must  be  constructed  for  the  pressure 
used.  If  the  air  is  to  be  eliminated  at  the  return  header  of  the  unit,  a 
high  pressure  air  valve  can  be  used ;  otherwise  the  air  may  be  passed  with 
the  condensate  through  the  high-pressure  return  trap,  with  some  danger 
of  return  pipe  corrosion  and  the  problem  of  its  elimination  at  some  other 
point  in  the  system. 

The  connections  for  steam  and  return  piping  to  unit  heaters  must 
always  be  calculated  on  the  basis  of  the  high  heat  emission  or  condensation 
rate  of  such  devices.  The  pipe-size  tables  given  in  Chapter  32  may  be 
used  for  unit  heater  work  by  multiplying  EDR  values  by  240  to  get  Btu 
values. 

OTHER  TYPES  OF  UNITS 
All  Electric 

The  foregoing  discussion  relates  generally  to  units  in  which  steam  or  hot 
water  is  used  as  the  heating  medium.  On  rare  occasions  electrical 
resistances  are  used  as  the  heating  element.  These  are  applied  only  where 
electric  power  is  abundant  and  cheap  and  where  other  forms  of  fuel  are 
scarce  and  expensive.  (See  Chapter  39.) 

Direct  Fired 

A  recent  development  in  gas  burning  equipment  is  the  direct-fired 
industrial  unit  heater.  These  heaters  are  of  the  warm  air  type  and  are 
equipped  with  fans  which  cause  the  air  to  pass  over  the  heating  surfaces 
at  a  fairly  high  velocity  and  then  direct  the  warm  air  in  to  the  space  to  be 
heated.  As  is  the  case  with  the  steam  fed  unit  heaters,  the  gas  fired 
appliances  may  be  used  for  heating  stores,  shops,  and  warehouses.  They 
usually  are  suspended  in  the  space  to  be  heated  and  in  most  instances 
leave  the  entire  floor  and  wall  area  free  for  commercial  use.  Partial  or 
complete  automatic  control  also  may  be  secured  on  appliances  of  this  type. 
This  type  of  heater  is  often  used  for  temporary  heat  during  building 
construction  or  where  the  installation  of  a  steam  or  hot  water  plant  is  for 
some  reason  not  justified. 

Turbine  Driven 

Where  high  pressure  steam  is  available  it  is  sometimes  used  to  drive  a 
steam  turbine  direct-connected  to  the  unit  heater.  The  exhaust  from 
this  turbine,  reduced  in  pressure,  is  then  passed  into  the  heating  coil 
where  it  is  condensed  and  returned  to  the  boiler. 

INDUSTRIAL  USES 

In  addition  to  their  prime  function  of  heating  buildings,  unit  heaters 
may  be  adapted  to  a  number  of  industrial  processes,  such  as  drying 

226 


CHAPTER  13 — UNIT  HEATERS,  VENTILATORS,  AND  COOLERS 

and  curing,  with  which  the  use  of  heated  air  in  rapid  circulation  with 
uniform  distribution  is  of  particular  advantage.  They  may  be  used  for 
moisture  absorption,  such  as  fog  removal  in  dye-houses,  or  for  the  pre- 
vention of  condensation  on  ceilings  or  other  cold  surfaces  of  buildings  in 
which  process  moisture  is  given  off.  When  such  conditions  are  severe,  it 
is  necessary  that  the  heaters  draw  air  from  outside  in  enough  volume  to 
provide  a  rapid  air  change  and  that  they  operate  in  conjunction  with 
ventilators  or  fans  for  exhausting  the  moisture-laden  air.  (See  discussion 
of  condensation  in  Chapter  7.) 

Information  on  the  control  of  unit  heaters  will  be  found  in  Chapter  14. 

UNIT  VENTILATORS2 

A  unit  ventilator  must  be  pleasing  in  design  because  it  is  generally 
used  where  it  must  harmonize  with  the  furniture  or  with  the  decorative 
scheme.  It  consists  usually  of  a  rectangular  steel  cabinet  finished  with  an 
enameled  surface  and  containing  the  following  necessary  or  optional 
parts: 

1.  Outside  air  inlet. 

2.  Inlet  damper  for  closing  the  opening  to  the  outside  air  inlet  when  the  unit  is  not 
in  use. 

3.  Adhesive  or  dry  type  filters  for  cleaning  the  air  (optional). 

4.  A  heating  element  usually  of  special  design  and  intended  for  low  pressure  steam. 

5.  Motor  and  fan  assembly. 

6.  Mixing  chamber  where  warm  and  cold  air  streams  are  brought  together.     (No 
mixing  chamber  is  normally  provided  where  sectional  type  con  vectors  are  used.) 

7.  Outdoor  air  inlet  and  recirculating  air  mixing  damper  (optional). 

8.  Device  for  ozonizing  air  (optional). 

9.  Discharge  grille  or  diffuser. 

10.  Temperature  control  arrangement. 

The  primary  functions  of  a  unit  ventilator  are: 

1.  To  supply  a  given  quantity  of  outdoor  air  for  ventilation  or  to  mix  indoor  and 
outdoor  air. 

2.  To  warm  the  air  to  approximately  the  room  temperature  if  the  unit  is  intended  for 
ventilation  only,  or  to  a  higher  temperature  if  it  is  intended  to  take  care  of  all  or  a  part 
of  the  heat  transmission  losses  from  the  room. 

3.  To  control  the  temperature  of  the  air  delivered  so  as  to  prevent  both  cold  drafts 
and  overheating. 

4.  To  deliver  air  to  the  room  in  such  a  manner  that  proper  distribution  is  obtained 
without  drafts. 

5.  To  recirculate  room  air  for  the  purpose  of  heating  or  promoting  comfort  when 
ventilation  is  unnecessary. 

6.  To  perform  all  its  functions  without  objectionable  noise. 

In  addition  to  these  functions,  unit  ventilators  frequently  are  arranged 
so  that  the  air  supplied  may  be  cleaned  by  means  of  filters  of  either  the 
dry  or  viscous  type.  If  filters  are  used,  the  proper  allowance  must  be 
made  for  the  Increased  resistance  offered  to  the  air  flow.  Humidifiers  in 
unit  ventilators  are  rather  difficult  to  control  and  are  only  furnished  upon 
special  order. 


*A  roof  ventilator  is  sometimes  termed  a  wiit  ventilator*.   For  information  on  roof  ventilators*  see 
Chapter  4. 

227 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

1.  Air  Supply  for  Ventilation.    The  outdoor  air  supply  for  ventilation 
is  delivered  by  motor-driven  fans  operated  at  comparatively  low  speeds, 
the  back  of  the  cabinet  being  connected  to  the  outside  through  rust-proof 
louvers  and  screens.    Air  quantities  may  be  estimated  on  the  basis  of  data 
given  in  Chapter  2.    (See  A.S.H.V.E.  Ventilation  Standards.) 

2.  Warming  Incoming  Air.    The  air  is  heated  by  passing  it  through 
specially  designed  convectors.     The  amount  of  heating  surface  to  be 
provided  in  the  unit  is  determined  by  the  volume  of  air  to  be  heated  and 
the  temperature  range.     If  the  unit  is  to  be  used  for  supplying  air  for 
ventilation  only,  the  convector  must  be  sufficient  in  capacity  to  maintain 
a  final  air  temperature  of  about  70  F.    If  the  unit  is  to  be  used  for  heating 
as  well  as  for  ventilation,  the  convector  must  be  sufficient  to  maintain  the 
necessary  final  air  temperature  for  the  conditions  involved. 

3.  Control  of  Temperature.    This  is  accomplished  by  varying  the  tem- 
perature of  the  air  discharged  from  the  unit  (1)  by  the  automatic  opera- 
tion of  a  mixing  damper  which  controls  the  relative  quantities  of  air 
being  blown  through  the  heating  unit  or  by-passed  around  it,  (2)  by 
operation  of  valves  on  different  layers  of  convector  surfaces,  or  (3)  by 
variation  in  the  temperature  of  the  circulating  heating  medium. 

The  outside  air  inlet  damper  and  recirculating  damper  (where  one  is 
provided)  should  be  so  connected  that  there  will  be  an  uninterrupted 
supply  of  air  to  the  fans  at  all  times  the  unit  is  in  operation.  These 
dampers  may  be  operated  by  hand  or  by  pneumatic  or  electric  motors 
manually  controlled  from  some  central  point. 

These  dampers  may  also  be  linked  together,  in  the  form  of  mixing 
dampers  and  be  controlled  by  a  thermostat  in  the  cold  air  intake,  by  a 
differential  thermostat  acted  upon  by  both  the  cold  air  and  the  recircu- 
lated  air,  or  by  a  thermostat  in  the  two  streams  of  air  after  they  are 
mixed,  so  as  to  keep  the  relative  proportion  of  air  taken  in  from  out-of- 
doors  commensurate  with  outside  temperatures  and  to  prevent  drafts  of 
cold  air  being  blown  through  the  unit  into  the  room. 

Provision  should  be  made  for  the  inlet  damper  to  close  automatically 
whenever  the  fans  are  shut  down,  and  not  to  open  until^  the  room  is 
properly  heated  when  the  fans  are  again  started.  The  minimum  tem- 
perature of  the  air  delivered  by  the  machine  should  be  regulated  auto- 
matically by  a  thermostat  in  the  outlet  air  which  controls  the  temperature 
of  the  heated  convector,  or  this  minimum  temperature  may  be  main- 
tained by  properly  mixing  the  inside  and  outside  air  by  means  of  the 
mixing  dampers  under  thermostatic  control  referred  to  above.  Another 
thermostat  in  the  recirculated  air  intake  to  the  unit  or  elsewhere  in  the 
room  controls  by-pass  dampers  or  the  supply  of  heating  medium,  or^both, 
so  as  to  control  the  temperature  of  the  air  leaving  the  unit  according  to 
the  heat  requirements  of  the  room.  In  addition  to  these  thermostats,  a 
room  thermostat  is  needed  to  control  any  other  heat  sources  for  the 
room.  (See  Chapter  14.) 

Thermostats  for  controlling  by-pass  dampers  must  ^be  of  the  inter- 
mediate type  to  hold  the  dampers  in  intermediate  positions  to  prevent 
objectionable  drafts.  When  direct  radiators  are  used  in  conjunction  with 
unit  ventilators,  the  control  is  usually  arranged  so  as  automatically  to 
open  the  valves  to  the  direct  radiators  when  the  room  temperature  falls 
about  2  deg  below  the  setting  of  the  thermostat  for  the  unit  ventilator. 

228 


CHAPTER  13 — UNIT  HEATERS,  VENTILATORS,  AND  COOLERS 

Another  arrangement  "opens  the  radiator  valve  whenever  the  unit  venti- 
lator control  reaches  the  full  heating  position.  Further  information  on 
this  subject  is  contained  in  Chapter  14. 

4.  Distribution.    This  function  is  governed  by  the  proper  selection  and 
location  of  the  unit.     Diffusion  and  distribution  are  dependent  upon  a 
relatively  high  velocity  air  stream  discharged  in  a  generally  vertical 
direction,  and  in  order  to  insure  satisfactory  diffusion  in  the  room  the  less 
the  difference  between  the  temperature  of  the  air  discharged  from  the  unit 
and  that  of  the  room  air,  the  better.    With  a  final  temperature  above 
110  F,  excessive  stratification  of  the  air  may  be  experienced.    Trouble- 
some drafts  may  be  eliminated  to  a  large  extent  if  a  static  pressure  is 
built  up  in  the  room. 

5.  Recirculation  of  air  requires  less  fuel  than  does  the  use  of  all  out- 
side air  and  aids  in  heating  up  quickly.     Certain  units  are  designed  to 
recirculate  all  air  at  all  times,  except  when  the  admission  of  outside  air  is 
needed  to  regulate  room  temperatures.     Under  this  arrangement,  the 
outside  air  for  ventilating  purposes  is  obtained  solely  from  infiltration,  but 
the  amount  thus  obtained  may  or  may  not  be  sufficient  to  meet  legal 
ventilating  requirements  for  public  buildings.    Recirculation  of  the  air  in 
schools   is   therefore  prohibited   by  ordinance  in   many   communities. 
Ventilating  systems  in  schools  should  be  arranged  for  taking  in  a  suf- 
ficient quantity  of  air  to  constitute,  with  infiltration,  not  less  than  10  cfm 
per  occupant  of  a  room. 

6.  Quiet  Operation.    Since  the  unit  ventilator  is  generally  set  in  close 
proximity  to  the  room  occupants,  it  must  operate  with  exceeding  quietness. 

SPLIT  AND  COMBINED  SYSTEMS 

In  a  split  system  the  unit  is  used  primarily  for  ventilation.  Air  is 
delivered  to  the  room  at  very  near  the  room  temperature,  and  enough 
separate  direct  heaters  are  placed  in  the  room  to  warm  it  to  the  desired 
temperature,  independently  of  the  unit.  Their  principal  advantage  lies 
in  offsetting  the  cooling  effect  of  window  and  wall  surfaces  long  before 
these  can  be  heated  to  room  temperature  and  in  retaining  heat  for  this 
purpose  after  the  ventilation  is  shut  down. 

Where  the  unit  ventilator  selected  has  a  capacity  more  than  sufficient 
to  warm  the  air  needed  to  meet  the  ventilating  requirements,  a  cor- 
responding reduction  may  be  made  in  the  amount  of  direct  heating  surface 
installed.  The  greater  the  amount  of  excess  capacity  of  the  unit,  the  more 
efficient  will  be  the  temperature  regulation  of  the  room.  The  split 
system  permits  the  heating  of  the  room  during  failure  of  electric  current, 
since  the  direct  radiators  will  furnish  ,,heat,  but  it  permits  a  careless 
operator  to  avoid  operating  the  "'ventilating  equipment. 

A  combined  system  employs  the  unit  ventilator  alone,  its  capacity  being 
sufficient  both  for  ventilation  and  for  supplying  the  heat  loss.  Direct 
heating  surface  is  omitted  altogether.  It  becomes  necessary  then  that  the 
fan  be  running  whenever  the  room  is  to  be  heated  but  this  also  gives 
assurance  of  ventilation,  especially  if  automatic  dampers  are  used  in  the 
air  intake  from  out-of-doors  and  in  the  recirculating  intake  arranged  so  as 
to  give  a  certain  quantity  of  air  from  the  outside  (commensurate  with 
weather  conditions)  whenever  the  unit  is  operating  and  after  the  room  is 

229 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

heated.  The  cost  of  installation  of  a  combined  system  is  usually  less  than 
that  of  a  split  system  and  there  is  less  danger  of  overheating,  but  if  the 
electric  energy  fails  there  will  be  practically  no  heating. 

LOCATION  OF  UNIT 

The  location  of  the  unit  ventilator  in  a  room  is  important.  Wherever 
possible  it  should  be  placed  against  an  outside  wall.  It  is  difficult  to 
obtain  proper  air  distribution  if  the  unit  is  erected  either  on  an  inside  wall 
or  in  a  corner  of  the  room.  Standard  units  discharge  the  air  stream  up- 
ward, but  for  special  cases  units  may  be  installed  to  discharge  air  hori- 
zontally. Units  may  be  set  away  from  the  wall  or  partially  recessed  into 
the  wall  to  save  space  without  materially  affecting  the  results.  The  air 
inlet  may  enter  the  cabinet  at  the  back  at  any  point  from  top  to  bottom. 

VENTS 

The  size  and  location  of  the  vent  outlet  is  important.  In  many  cases 
the  sizes  for  public  buildings  are  regulated  by  law,  but  the  location  of  the 
vents  generally  is  left  to  the  discretion  of  the  engineer. 

Best  results  have  been  obtained  with  a  velocity  through  the  vent 
openings  nearly  equal  to  that  at  which  the  air  is  introduced  into  the  room, 
thus  maintaining  a  slight  pressure  in  the  room.  Calculated  velocities  at 
the  vent  openings  of  from  600  to  800  fpm  produce  the  best  diffusion  results 
from  this  system. 

The  cross-sectional  area  of  the  vent  flue  itself  may  be  figured  on  the 
basis  of  15  sq  in.  of  flue  for  each  100  cfm.  Thus  the  vent  flue  area  of  a 
flue  for  a  room  equipped  with  one  1200  cfm  unit  ventilating  machine 
would  be  180  sq  in.  The  area  of  vent  flue  opening  from  the  room  may  be 
figured  on  the  basis, of  25  sq  in.  per  100  cfm, 

In  school  buildings  provided  with  wardrobes  or  cloakrooms  the  vents 
may  be  so  located  that  the  air  shall  pass  through  these  spaces,  heating  and 
ventilating  them  with  air  which  otherwise  would  be  passed  to  the  outside 
without  being  used, to  the  best  advantage.  Many  state  codes  for  venti- 
lation of  public  buildings  make  this  arrangement  mandatory. 

There  has  been  much  controversy  over  the  use  of  corridor  ventilation 
in  school  building  practice,  one  group  holding  the  view  that  when  each 
classroom  has  a  separate  vent  flue  there  is  a  minimum  fire  risk  and  less 
likelihood  of  cross-contamination,  while  others  emphasize  the  economy 
features  of  the  corridor  discharge  and  minimize  the  fire,  contamination, 
and  other  hazards. 

CAPACITIES 

Unit  ventilators  are  available  in  air  capacities  ranging  from  450  cfm  to 
6000  cfm  and  with  corresponding  heat  capacities  (above,  that  required  for 
ventilation  purposes  based  upon  an  outside  temperature  of  zero  and  an 
inside  temperature  of  70  F)  ranging  from  30  Mbh  to  144  Mbh  (1  Mbh  = 
1000  Btu  per  hour).  Some  manufacturers  furnish  a  unit  with  several 
heating  capacities  for  each  air  capacity,  thus  enabling  the  -engineer  to 
select  the  unit  best  adapted  to  the  heating  and  ventilating  load.  Capaci- 
ties should  be  determined  in  accordance  with  the  A.S.H.V.E.  Staadard 

230 


CHAPTER  13 — UNIT  HEATERS,  VENTILATORS,  AND  COOLERS 

Code  for  Testing  and  Rating  Steam  Unit  Ventilators3.  Typical  capacities 
are  given  in  Table  3. 

The  amount  of  heat  to  be  supplied  by  the  unit  ventilator  will  depend  on 
the  amount  of  air  passed  through  the  unit  and  the  temperature  range 
through  which  the  air  is  heated.  The  weight  of  air  (W)  to  be  circulated 
per  hour  is  fixed  by  the  ventilating  requirements. 

If  no  direct  heating  surface  (radiation)  is  installed,  the  combined 
heating  and  ventilating  requirements  must  be  taken  care  of  by  the  unit 
ventilators,  and  the  total  heat  to  be  supplied  is  obtained  by  means  of  the 
following  formulae: 

When  all  of  the  air  handled  by  the  unit  is  taken  from  the  outside, 

Ht  =  0.24  W  (ty  -  to)  (1) 

W  =  dQ  (2) 

H  (3) 


where 


Q.24W 


d  =  density  of  air,  pounds  per  cubic  foot. 
H  —  heat  loss  of  room,  Btu  per  hour. 

Hv  =  heat  required  to  warm  air  for  ventilation,  Btu  per  hour. 
Ht  —  total  heat  requirements  for  both  heating  and  ventilation,   Btu  per  hour 

=  H  +  Hv. 

Q  =  volume  of  air  handled  by  the  ventilating  equipment;  cubic  feet  per  hour. 
t  —  temperature  to  be  maintained  in  the  room. 
t0  —  outside  temperature. 
ty  =  temperature  of  the  air  leaving  the  unit. 
W  —  weight  of  air  circulated,  pounds  per  hour. 
0.24  =  specific  heat  of  air  at  constant  pressure. 

From  Equations  1,  2  and  3: 

-fc)  (4) 


Example  1 .  The  heat  loss  of  a  certain  room  is  24,000  Btu  per  hour,  and  the  ventilating 
requirements  are  1000  cfm.  If  the  room  temperature  is  to  be  70  F  and  all  air  is  taken 
from  the  outside  at  zero,  what  will  be  the  total  heat  demand  on  the  unit  if  it  is  required 
to  provide  for  both  the  heating  and  ventilating  requirements  (combined  system)? 

Solution.    H  =  24,000;  d  =  0.075 

Q  =  1000  x  60  =  60,000  cfh;  t  «  70  F;  tQ  =  0  F. 

Substituting  in  Equation  4: 

Ht  =  24,000  +  0.24  x  0.075  x  60,000  (70-0)  =  99,600  Btu 
. 24,000 


0.24  x  0.075  x  60,000 


70  -  92.2  F 


When  part  of  the  air  handled  by  the  unit  is  taken  from  the  room  and  the 
remainder  from  the  outside, 


Ht  =  0.24W0  &-*>)+  O-24  wi  (h  -  *)  (5) 


•Adopted  1032.    See  AJ5.H.V.R.  TRANSACTZO??^  Vol.  38, 

2S1 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


where 


=  weight  of  air,  pounds  per  hour  taken  from  out-of-doors. 
=  weight  of  air,  pounds  per  hour  taken  from  the  room. 


(6) 

(7) 


where 


Qo 

a 


=  density  of  air,  pounds  per  cubic  foot  at  temperature  to 
=  density  of  air,  pounds  per  cubic  foot  at  temperature  t. 
=  volume  of  air  taken  in  from  the  outside,  cu  ft  per  hr. 
=  volume  of  air  taken  in  from  the  room,  cu  ft  per  hr. 


H 


+ 


0.24  (Wo  + 
H  +  0.24  d0  Qo  (t  - 


(8) 
(9) 


Equations  5,  6,  7,  8,  and  9  may  be  used  in  the  same  manner  as  is 
illustrated  above  for  Equations  1,  2,  3,  and  4.  It  may  be  noted  in  Equa- 
tion 9,  representing  the  total  heat  requirements,  that  as  the  quantity 
Qo  is  diminished  the  heat  requirements  for  the  unit  diminish  very 
materially. 

In  Example  1,  if  the  quantity  of  air  taken  in  from  the  outside  is  reduced 
to  zero,  or  all  of  the  air  handled  by  the  unit  is  recirculated,  the  total  heat 
requirements  Ht  reduce  from  99,600  Btu  to  24,000  Btu,  or  to  about  one 
fourth.  Such  a  unit  handling  one-third  of  its  air  volume  from  the  outside 
and  two  thirds  from  the  room  would  show  a  total  heat  requirement  of 

24,000  +  99>6Q°  7"  24'°QQ  «  59,200  Btu.  Units  designed  and  operated 
o 

on  this  principle  show  an  average  heat  requirement  and,  therefore,  a  boiler 
capacity  requirement  of  less  than  50  per  cent  of  that  required  for  units 
taking  all  their  air  from  the  outside. 

If  all  of  the  air  is  recirculated,  the  total  heat  required  is  the  same  as  the 
heat  loss  of  the  room,  or 


0.24  W  (ty  -  0 


TABLE  3.    TYPICAL  CAPACITIES  OF  UNIT  VENTILATORS  FOR 
AN  ENTERING  AIR  TEMPERATURE  OF  ZERO 


(10) 


TOTAL  CAPACITY  IN  SQUARE  FEET 

CAPACIT?  AVAILABLE  FOR  HEAT- 

OF EQUIVALENT  DIRECT  HEATING 

ING  THE  ROOM  IN  SQUARE  FEET 

CUBIC  FEET  OF 

SURFACE  (RADIATION) 

OF  EQUIVALENT  DIRECT  HEATING 

FINAL  AIR  TEMPERA- 

AIR PER  MINUTE 

SURFACE  (RADIATION) 

TURE  (DEG  FA.HR) 

EDR 

Mbh 

EDR 

Mbh 

600 

285 

68 

95 

23 

105 

750 

350 

84 

115 

28 

105 

1000 

455 

110 

150 

36 

105 

1200 

565 

136 

190 

46 

105 

1500 

705 

169 

235 

56 

105 

232 


CHAPTER  13 — UNIT  HEATERS,  VENTILATORS,  AND  COOLERS 

If  the  heat  loss  of  the  room  is  to  be  taken  care  of  by  the  direct  heating 
surface,  the  unit  ventilators  will  be  required  to  warm  the  air  introduced 
for  the  ventilating  requirements.  Therefore: 

Hv  =  0.24  W  (ty  -  t0)  (11) 

In  this  case  t?  should  be  equal  to  or  slightly  higher  than  i.  If  the  unit 
ventilator  were  of  such  capacity  as  to  exactly  provide  for  the  ventilating 
requirements,  the  direct  radiation  would  be  selected  on  the  usual  basis. 
However,  it  is  necessary  to  employ  a  unit  which  may  not  exactly  meet  the 
ventilating  requirements,  since  standard  units  are  usually  rated  in  terms 
of  the  volume  of  air  that  will  be  delivered  at  a  certain  temperature  ty  for 
an  initial  temperature  of  tQ.  Therefore  a  certain  amount  of  heat  (flh) 
may  be  available  from  the  unit  ventilator  for  heating  purposes,  as  pre- 
viously stated,  and  the  amount  of  equivalent  direct  heating  surface  may, 
if  desired,  be  deducted  from  the  amount  required  for  heating  the  room. 

ATTIC  FANS 

Attic  fans,  used  during  the  warm  months  of  the  year  to  draw  large 
volumes  of  outside  air  through  a  house,  offer  a  means  of  using  the  com- 
parative coolness  of  outside  evening  and  night  air  to  bring  down  the 
inside  temperature  of  a  house. 

Because  the  low  static  pressures  involved  are  usually  less  than  Y%  in.  of 
water,  disc  or  propeller  fans  are  generally  used  instead  of  the  blower  or 
housed  types.  The  fans  should  have  quiet  operating  characteristics,  and 
they  should  be  capable  of  giving  about  thirty  air  changes  per  hour.  The 
two  general  types  of  attic  fan  installations  in  common  use  are: 

Open  attic  fans,  in  which  the  fan  is  installed  in  a  gable  or  dormer  and 
one  or  more  grilles  are  provided  in  the  ceilings  of  the  rooms  below. 
Fresh  air,  which  enters  the  house  through  open  windows,  is  drawn  into 
the  attic  through  the  grilles,  and  is  discharged  out-of-doors  by  the  fan. 
An  attic  stairway  may  be  used  in  place  of  the  central  grille.  It  is 
essential  that  the  roof  and  the  attic  walls  be  free  from  air  leaks. 

Boxed-infan,  in  which  the  fan  is  installed  within  the  attic  in  a  box  or 
housing  directly  over  a  central  ceiling  grille,  or  in  a  bulkhead  enclosing 
an  attic  stair.  The  fan  may  be  connected  by  a  duct  system  to  the 
grilles  in  individual  rooms.  Fresh  air  entering  through  the  windows  of 
the  rooms  below  is  discharged  into  the  attic  space  and  escapes  to  the 
outside  through  louvers,  dormer  windows,  or  screened  openings  under 
the  eaves. 

The  locations  of  the  fan,  the  outlet  openings,  and  the  grilles  should  be 
chosen  after  consideration  of  the  room  and  attic  arrangement  in  order  to 
give  uniform  air  distribution  in  the  individual  rooms  served.  If  the  outlet 
for  the  air  is  not  on  the  side  away  from  the  direction  of  the  prevailing 
wind,  openings  should  be  provided  on  all  sides.  Kitchens  should  be 
separately  ventilated  because  of  the  fire  hazard,  and  to  prevent  the 
spread  of  cooking  odors. 

The  operating  routine  which  will  secure  best  results  with  an  attic  fan  is 
an  important  consideration.  A  typical  routine  might  require  that  in  the 
late  afternoon  when  the  outdoor  temperature  begins  to  fall,  the  windows 

233 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

on  the  first  floor  and  the  grilles  in  the  ceiling  or  the  attic  floor  should  be 
opened,  and  the  second  story  windows  should  be  kept  closed.  This  will 
place  the  principal  cooling  effect  in  the  living  rooms.  Shortly  before 
bedtime,  the  first  floor  windows  may  be  closed  and  those  on  the  second 
floor  opened,  to  transfer  the  cooling  effect  to  the  sleeping  rooms.  A  time 
clock  may  shut  off  the  fan  before  waking  time,  or  the  fan  may  be  stopped 
manually  at  a  later  hour. 

A  disadvantage  arising  from  the  passing  of  a  great  amount  of  outside 
air  through  a  house  is  the  dust  nuisance,  which  varies  considerably  in 
different  locations.  Persons  suffering  from  allergic  diseases  caused  by  air- 
borne pollens  will  have  their  troubles  increased  with  attic  type  coolers. 

Some  typical  data  on  an  attic  fan  installation  in  an  average  six-room 
house  of  frame  construction  containing  14,000  cu  ft  and  located  in  the 
southern  part  of  this  country  are : 


Installation  cost.. 
Fan  data 


Operating  period.. 


Power  consumption 


$75  to  $400,  average  $250 

9000  cfm  average,  280  rpm  if  belt  driven,  570  rpm  if  direct 
connected,  500  watts  input 

April  15  to  October  15,  intermittently  as  weather  con- 
ditions demand 

500  kwh  per  year  for  8  months'  operation 


UNIT  COOLERS 

A  unit  cooler,  as  defined  in  Chapter  41,  is  a  device  usually  comprising 
an  extended-surface  element  and  a  motor-driven  fan  mounted  integrally 
in  a  housing,  suitable  to  be  placed  within  or  adjacent  to  the  room  served. 
The  refrigerating  medium  is  brought  to  the  unit  from  an  outside  source, 
and  the  fan  drives  air  over  the  cooling  element;  generally,  no  d,ucts  are 
attached  to  inlet  or  outlet.  With  provision  for  filtering  the  air  and  taking 
in  outdoor  air  for  ventilation,  the  apparatus  becomes  a  unit  conditioner 
(Chapter  12).  An  alternative  design  uses  chilled  water  or  brine  spray  for 
cooling  the  air;  it  is  essentially  a  small  compact  air  washer  with  built-in 
fan  and  accessory  equipment. 

The  principal  field  for  unit  coolers  is  in  cold-storage  plants,  fur-storage 
vaults,  packing  houses,  provision  stores,  brewery  fermentation  and  stock 
rooms,  and  industrial  process  work.  Coolers  have,  to  a  considerable 
extent,  supplanted  the  bunker  coils  heretofore  placed  on  ceilings  and  walls, 
because  of  demonstrated  advantages  with  respect  to :  compactness,  first 
cost,  maintenance  expense,  damage  from  drips,  ease  of  defrosting,  main- 
tenance of  sanitary  conditions,  uniformity  of  temperature  throughout  the 
space  served,  and  uniformity  of  temperature  under  variable  load  con- 
ditions, as  well  as  control  of  humidity  and  circulation  of  room  air  when 
conducive  to  improved  results. 

A  typical  suspended  unit  is  shown  in  Figs.  3  and  4.  A  motor-driven 
propeller-type  fan  is  bracketed  to  the  frame  of  a  sheet-metal  housing  that 
contains  an  extended-surface  coil,  and  a  double  set  of  louvers  acting  also 
as  a  moisture  eliminator  is  provided  at  the  outlet  side.  The  horizontal 
louvers  are  adjustable  to  direct  the  air  downward,  horizontally,  or  upward, 
as  desired.  The  lower  part  of  the  housing  forms  a  drip  pan,  requiring  a 
drain  connection  to  dispose  of  the  condensation  when  dehumidifying  air 

234 


CHAPTER  13 — UNIT  HEATERS,  VENTILATORS,  AND  COOLERS 

at  usual  room  temperatures,  or  of  the  water  when  defrosting  in  low- 
temperature  service.  A  cabinet-type  unit  for  floor  mounting  is  shown  in 
Fig.  5 ;  other  designs  are  illustrated  in  the  Catalog  Data  Section  at  the  rear 
of  this  volume. 


«  ^  — i 

PI— Hon^r 

I    I „  Water       I 


Connection 


Louvres 


Connection 


FIG.  3.  CEILING  UNIT 


l/ertfca/   Diffusing 
Eliminators 


Front 
On   L 


%  Orif> 
Conn 


FIG.  4.  ELEVATION  THROUGH  LINE  AA 


Depending  upon  the  arrangement  of  the  cooling  coil,  chilled  water, 
brine,  or  a  direct-expansion  refrigerant  may  be  employed.  For  cooling 
service  at  or  near  ordinary  room  temperatures,  the  considerations 
affecting  a  choice  of  cooling  medium  are  those  discussed  in  Chapter  12  for 
unit  air  conditioners.  At  lower  temperatures,  as  for  cold-storage,  the 

235 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

refrigerant  system  is  usually  dictated  by  the  requirements  of  other 
refrigeration  services  supplied  from  the  same  condensing  unit  or  from  a 
central  plant. 

Details  of  construction  employed  in  unit  coolers  are  generally  similar  to 
those  for  unit  air  conditioners,  with  special  attention  paid  to  the  use  of 
non-corroding  materials.  Temperature  control  is  obtained  by  starting 
and  stopping  the  fan,  with  or  without  regulation  of  the  cooling  liquid  or 
direct-expansion  refrigerant  admitted  to  the  coil.  Usually,  a  thermostatic 
control  is  provided  ahead  of  the  expansion  valve  at  the  inlet  to  the  coil, 
tending  to  maintain  constant  temperature  and  pressure  inside  the  coil 
regardless  of  cooling  load,  with  a  float  at  the  outlet  to  prevent  accumu- 
lation of  liquid  refrigerant  in  amounts  sufficient  to  interfere  with  dis- 


1  til 


1 

fr"^"f. 

^•••T""K 

~Tr'"2~Kt" 

~jH~  if" 

1  1  I 

7          I  i  u 

(1 

i< 

J  ,              Coo///7y 

H     «      £—  =  --  =  -- 

£"/e/77en?-p>           | 

! 

!'  ii   e--=.-=-  = 

ii  *_,j  

!'     '  i 

aai""V* 

II  j  1 

i  i^  ;i 

J 

ii    !  /       r/oor-i? 

rtp  Pon                ^  , 

/v// 

V  V  ////////////// 

s/s////////  ////?/// 

/// 

FIG.  5.    CABINET  TYPE  COOLING  UNIT 

tribution  between  the  various  unit  coolers  served  by  a  central  condensing 
unit. 

Ratings  of  unit  coolers  may  be  expressed  in  Btu  per  hour  or  in  tons  of 
refrigeration,  with  specified  quantity,  temperature,  and  humidity  of  air 
at  the  inlet,  and  with  a  stipulated  pressure  or  temperature  maintained 
within  the  cooling  coil  when  using  direct-expansion  refrigerants.  When 
chilled  water  or  brine  are  used  as  the  cooling  media,  the  quantity  and 
inlet  temperature  must  be  given.  Ratings  and  dimensions  of  representa- 
tive makes  of  unit  coolers  are  given  in  the  Catalog  Data  Section. 


PROBLEMS  IX  PRACTICE 

1  •  Is  it  better  to  use  high  pressure  or  low  pressure  steam  in  unit  heaters? 

The  answer  to  this  question  depends  upon  the  following  circumstances:  If  steam  is  used 
only  for  heating  purposes,  it  is  usually  best  to  design  the  entire  system  for  low  pressure 
steam.  When  steam  is  generated  at  high  pressure  for  other  purposes,  it  can  be  used 
either  at  full  boiler  pressure  or  at  reduced  pressure  in  the  unit  heaters.  If  the  steam 
pressure  is  reduced,  the  heating  elements  should  be  capable  of  withstanding  the  full  boiler 
pressure.  When  steam  at  full  boiler  pressure  is  used  in  the  heating  elements,  the  heating 
surface  should  be  reduced  so  that  the  outlet  temperature  will  not  be  more  than  70  F 
higher  than  the  inlet  temperature.  Wjth  the  use  of  high  pressure  steam  special  care  must 
be  exercised  in  venting  the  units  of  air,  in  preventing  flash  steam  in  the  returns,  and  in 
preventing  corrosion  from  superheated  returns. 

236 


CHAPTER  13 — UNIT  HEATERS,  VENTILATORS,  AND  COOLERS 

2  •  How  should  heat  losses  he  calculated  for  a  huilding  using  unit  heaters? 

The  heat  losses  should  be  calculated  in  exactly  the  same  manner  as  for  any  other  heating 
system.  If  the  method  of  calculation  takes  into  consideration  the  variation  in  tem- 
perature from  the  floor  to  the  ceiling,  the  temperature  variation  should  be  reduced  when 
calculating  the  heat  losses  for  a  unit  heater  job.  This  is  advisable  because  with  unit 
heaters  the  temperature  variation  between  the  floor  and  the  ceiling  is  from  }•£  to  1  F 
per  foot  of  elevation,  whereas  with  direct  radiators  or  pipe  coils,  this  variation  may  be 
twice  as  great.  Unless  the  ceiling  height  is  more  than  15  ft,  the  temperature  variation 
between  the  floor  and  the  ceiling  is  usually  neglected  when  unit  heaters  are  used. 

3  •  On  what  hasis  should  unit  heaters  he  selected? 

Unit  heaters  should  be  selected  to  furnish  enough  heat  to  offset  the  heat  losses  and  to 
circulate  the  air  in  the  room  fast  enough  to  provide  good  heat  distribution.  In  the 
average  building,  if  the  outlet  temperature  does  not  exceed  the  inlet  temperature  by  more 
than  70  F,  sufficient  air  capacity  will  usually  be  provided  for  proper  circulation  if  the 
units  are  selected  strictly  on  the  basis  of  heating  capacity.  However,  if  the  units  are 
hung  unusually  high  or  if  the  heat  loss  is  low  in  proportion  to  the  volume  of  the  room, 
then,  in  order  to  obtain  the  desired  air  capacity,  it  is  usually  necessary  to  employ  more 
heaters  than  are  required  to  offset  the  normal  heat  loss.  Inasmuch  as  the  heat  distri- 
bution depends  upon  the  outlet  temperature,  the  outlet  velocity,  the  character  of  air  flow 
from  the  heater,  the  height  at  which  the  heaters  are  hung,  and  the  size  of  the  heater 
itself,  the  manufacturers'  literature  should  be  carefully  studied  in  determining  the  exact 
number  of  heaters  to  be  employed. 

4  •  Is  it  satisfactory  to  use  superheated  steam  in  unit  heaters? 

Superheated  steam  can  be  satisfactorily  used  in  unit  heaters  provided  the  capacity  is 
based  on  the  saturated  steam  temperature  and  not  on  the  total  temperature.  If  un- 
usually high  superheat  is  used,  trouble  may  be  experienced  from  the  excessive  expansion 
and  contraction  of  the  heating  elements. 

5  •  Is  it  satisfactory  to  install  one  unit  heater   as   the  total  load  on  a  coal 
fired  hoiler? 

Such  an  arrangement  is  impractical  if  the  unit  heater  is  started  and  stopped  in  keeping 
with  the  room  temperature.  However,  if  the  room  temperature  controls  the  steam  pres- 
sure and  the  unit  heater  is  arranged  to  start  when  there  is  steam  in  the  mains  and  to 
stop  when  there  is  no  steam  in  the  mains,  such  an  installation  will  be  satisfactory. 

6  •  Will  a  unit  heater  with  a  slow  speed  fan  he  more  quiet  than  one  with  a 
high  speed  fan? 

The  speed  of  the  fan  is  no  indication  of  quietness.  Quietness  is  a  function  of  the  type, 
diameter,  blade  form,  speed,  and  location  of  the  fan. 

7  •  Is  it  satisfactory  to  use  steam  at  pressures  less  than  atmospheric  for  unit 
heaters? 

If  the  air  inlet  temperature  is  above  freezing,  steam  at  any  pressure  may  be  used  in  the 
unit  heater.  If  the  inlet  temperature  is  below  freezing,  steam  of  at  least  5  Ib  pressure 
(or  with  a  positive  5  Ib  pressure  differential  between  supply  and  return)  should  be  used, 
and  the  steam  supply  should  never  be  throttled  or  the  heating  element  may  be  frozen. 

8  •  In  general,  what  is  the  primary  function  of  a  unit  ventilator? 

To  maintain  the  desired  room  air  conditions  as  to  temperature,  air  change,  and  air 
cleanliness,  without  drafts  regardless  of  variations  in  outdoor  temperature,  occupancy, 
sun  heat,  and  wind. 

9  •  What  are  the  usual  working  parts  of  a  unit  ventilator? 

A  fan  and  motor  assembly,  a  set  of  heating  elements,  outdoor  and  indoor  air  dampers. 
filters,  outlet  grille,  some  method  of  controlling  the  outlet  temperature  above  a  minimum 
of  60  F,  and  some  method  of  varying  the  outlet  temperature  in  keeping  with  the  room 
requirements.  All  of  these  parts  are  usually  enclosed  in  an  attractive  steel  cabinet  in 
which  the  piping  is  concealed. 

237 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

10  •  Do  all  unit  ventilators  introduce  a  constant  amount  of  outdoor  air? 

Certain  types  employ  full  recirculation  except  when  outdoor  air  is  obtained  by  throttling 
the  steam  valve  on  the  heating  element  so  the  proportion  of  outdoor  air  to  room  air  is 
varied.  This  is  a  very  economical  type  of  unit  ventilator  but  in  some  communities  it 
cannot  be  used  because  of  existing  laws  which  require  that  some  fixed  amount  of  outdoor 
air  be  introduced  whenever  the  room  is  occupied.  Certain  types  of  units  are  designed  to 
always  take  in  a  minimum  quantity  of  air  from  the  outside  and  to  automatically  vary 
this  with  the  weather. 

11  •  Where  should  a  unit  ventilator  he  located? 

In  the  center  of  the  longest  outside  wall  under  the  windows. 

12  •  What  further  precaution  should  he  taken  in  locating  unit  ventilators? 

With  most  unit  ventilators,  a  high  velocity  jet  of  air  is  discharged  toward  the  ceiling  at  a 
slight  pitch  toward  the  room ;  all  unit  ventilators  should  be  installed  in  such  manner  that 
this  jet  is  not  interfered  with.  For  this  reason  the  air  should  be  distributed  on  a  flat 
ceiling  without  beams,  but  if  beams  are  present,  the  unit  ventilator  should  be  so  located 
that  the  air  will  be  discharged  parallel  to  the  beams. 

13  •  When  unit  ventilators  are  installed  to  employ  variable  recirculation,  what 
special  precautions  are  necessary? 

Where  partial  recirculation  is  employed,  some  effective  means  should  be  installed  within 
the  cabinet  of  the  unit  ventilator  to  prevent  unheated  outdoor  air  from  being  blown  into 
the  room  through  the  room  air  opening  while  the  unit  is  mixing  indoor  and  outdoor  air. 
This  means  may  be  self-operating  dampers  placed  in  the  path  of  ^the  room  air,  or  filters 
so  arranged  that  the  outdoor  air  must  pass  through  them  before  it  can  enter  the  room. 

14  •  Generally  speaking,  should  direct  radiators  be  used  in  addition  to  unit 
ventilators  in  school  classrooms? 

The  best  practice  in  schoolrooms  is  to  place  as  much  heating  capacity  as  possible  in  the 
unit  ventilator  itself.  However,  in  selecting  the  unit  ventilator,  the  outlet  temperature 
should  not  exceed  110  F  and  the  rate  of  air  circulation  should  not  exceed  9  room  volumes 
per  hour  (anemometer  measurement)  or  7^4  room  volumes  per  hour  (A.S.H.V.E.  Code 
measurement).  If  the  heating  capacity  under  these  conditions  is  sufficient  to  heat  the 
room,  no  additional  radiation  is  required.  If  the  heating  capacity  is  not  sufficient,  direct 
radiation  should  be  used  to  make  up  the  required  total.  Radiators  always  tend  to  offset 
the  chilling  effect  of  cold  walls  and  windows  quicker  than  warm  air  does. 

15  •  Are  vent  outlets  required  with  unit  ventilators? 

Though  experience  has  indicated  that  in  practically  all  school  and  office  buildings  the 
cracks  around  the  windows,  doors,  and  baseboards  are  so  numerous  that  vents  are  not 
required,  in  many  communities  vents  are  required  by  existing  laws.  In  some  cases  the 
sizes  are  also  stipulated  in  the  laws.  When  the  size  is  not  stipulated,  vents  should  be 
designed  on  the  basis  of  a  velocity  not  greater  than  600  ft  per  minute.  Vent  flues  should 
always  be  provided  with  a  damper  in  order  that  they  may  be  throttled. 


238 


Chapter  14 

AUTOMATIC  CONTROL 

Apparatus  Sensitive  to  Temperature,  Apparatus  Sensitive  to 
Relative  Humidity,  Apparatus  Sensitive  to  Pressure,  Accessory- 
Apparatus,  Temperature  Control  Systems,  Control  of  Automatic 
Fuel  Appliances,  Individual  Room  Control,  Zone  Control,  In- 
dustrial Processes,  Air  Conditioning  Systems,  Seasonal  Operation 

\  UTOMATIC  controls  can  be  installed  on  any  type  of  heating, 
J~\.  ventilating,  or  air  conditioning  system  to  maintain  desired  con- 
ditions automatically,  and  with  maximum  operating  economy.  The 
variety  of  automatic  control  equipment  available  is  such  that  a  suitable 
control  system  can  be  devised  without  difficulty,  provided  that  the  con- 
ditions to  be  maintained  are  known  and  the  control  equipment  is  properly 
chosen.  This  chapter  outlines  briefly  the  various  types  of  control  appar- 
atus and  indicates  the  method  of*  their  application  to  typical  heating, 
ventilating,  and  air  conditioning  systems.  Specific  control  devices  and 
systems  are  described  in  the  Catalog  Data  Section  of  THE  GUIDE. 

Controls  are  applied  for  the  following  reasons: 

1.  To  maintain  conditions  required  for  human  comfort  and  efficiency. 

2.  To  maintain  conditions  required  for  industrial  processes. 

3.  To  obtain  economy  in  operation. 

4.  To  provide  necessary  safety  measures. 

CONTROL  APPARATUS 

The  various  pieces  of  control  apparatus  may  be  grouped  under  the 
following  general  headings: 

Apparatus  Sensitive  to  Temperature 

Temperature-sensitive  devices  which  will  respond  to  changes  in  tem- 
perature, and  which  will  motivate  equipment  to  compensate  for  the 
changes,  are  usually  called  thermostats.  They  have  many  specialized 
forms  for  use  in  specific  control  applications.  Thermostats  are  the 
detectors  of  a  control  system  which  identify  changes  in  desired  tempera- 
ture conditions  and  automatically  call  for  compensating  action. 

Thermostats  are  actuated  by  various  means,  all  of  which  have  the 
common  characteristic  of  responsiveness  to  small  changes  of  temperature. 
The  actuating  element  may  be  a  piece  of  bi-metal  in  straight,  helical,  or 
spiral  form  (Fig.  1),  which,  by  bending  slightly  as  the  temperature 
changes,  actuates  an  electric  or  pneumatic  switch  to  govern  the  controlled 
apparatus;  or  the  actuator  may  be  a  diaphragm,  bellows,  or  tube  filled 

239 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

with  a  volatile  liquid  (Fig.  2)  in  such  way  that  expansion  and  contraction 
with  changes  in  temperature  will  operate  the  controlled  apparatus  by  a 
direct  mechanical,  electric,  or  pneumatic  connection. 

A  room  or  watt  thermostat  in  its  simplest  form  contains  a  single  tempera- 
ture-sensitive element  which  is  so  set  that  it  maintains,  by  actuating  the 
controlled  system,  a  single  temperature.  The  two-temperature  or  dual 
thermostat  has  two  temperature-sensitive  elements,  one  of  which  is  set  for 
a  higher  temperature  than  the  other.  Such  a  thermostat  is  used  on  day- 
night  systems  where  the  night  temperature  is  to  be  lower  than  that 


STBAIGHT  3TK/P 


b.    Spiral  Type 


a.    Straight  Strip  Type 


c.    Curved  Strip  Type 


FIG.  1.    TYPICAL  BI-METALLIC  THERMOSTATIC  ELEMENTS 


Volatile  Liquid 


FIG.  2.    DIAPHRAGM  TYPE  THERMOSTAT 

maintained  during  the  daytime  hours.  Switching  the  control  from  one 
element  to  the  other  is  accomplished  by  an  external  or  an  internal  switch, 
which  can  be  operated  manually  or  by  a  time  device. 

Duct  type  thermostats  are  used  in  systems  where  the  equipment  must 
respond  to  changes  in  the  temperature  of  the  air  passing  through  a  duct. 
In  their  usual  form,  these  thermostats  are  so  constructed  that  their 
switching  mechanism  is  outside  the  duct,  while  the  temperature-sensitive 
element  projects  inside  into  the  air  stream. 

Thermostats  which  operate  in  liquids  have  the  same  general  construc- 
tion as  duct  thermostats  except  that  the  sensitive  element  is  usually 
enclosed  in  a  tube  to  keep  it  from  direct  contact  with  the  liquid.  They 

240 


CHAPTER  14 — AUTOMATIC  CONTROL 


are  used  in  pipes,  vats,  and  tanks,  and  are  called  immersion  thermostats. 
Such  a  thermostat  is  Illustrated  in  Fig.  3. 

Sometimes  surface  thermostats  are  used  in  place  of  duct  or  immersion 
thermostats.  These  devices,  so  constructed  as  to  respond  to  changes  in 
temperature  of  the  surface  of  the  duct  or  vessel  containing  a  fluid,  are 
clamped  or  screwed  to  such  surfaces  in  a  manner  which  will  provide  as 
rapid  as  possible  heat  transfer  between  the  surface  and  the  sensitive 
element. 

Apparatus  Sensitive  to  Relative  Humidity 

Devices  which  are  responsive  to  changes  in  the  relative  humidity  of  the 
surrounding  air,  and  which  will  motivate  equipment  to  compensate  for 
the  changes,  are  called  humidistats  or  hygrostats.  These  may  vary  con- 
siderably in  their  sensitive  elements,  but  they  all  operate  through  con- 
necting equipment  which  automatically  causes  humidifying  apparatus  to 
supply  more  or  less  moisture  as  required.  Some  of  the  more  complicated 


VACUUM 

RELEASE  AT  AWCUUtt 
GREATER  THAU  THAT 
CAUSED  BY!  JW  POOP 
THERTIOSTATIC  TRAP 


-RETURN  TO 
VACUUM  PUMP 


Fic.  3.  SELF-CONTAINED  THERMOSTAT  ON  HOT  WATER  TANK  WITH  VACUUM  RETURN 


ones  contain  essentially  two  thermostats,  one  working  on  a  dry-bulb 
temperature  and  the  other  on  a  wet-bulb  temperature  ;  by  proper  inter- 
connection of  the  parts  they  operate  to  maintain  a  definite  relation  be- 
tween these  two  temperatures.  Other  devices  use  elements,  directly 
sensitive  to  humidity,  made  of  special  wooden  blocks,  human  hair,  fiber, 
membranes,  or  strips  of  prepared  paper.  Hygrostats  are  available  for 
use  with  both  electric  and  pneumatic  control  systems. 

Apparatus  Sensitive  to  Pressure 

Use  is  made  of  devices  which  are  responsive  to  changes  in  pressure,  and 
which  will  motivate  equipment  to  compensate  for  the  changes.  Such 
devices  usually  depend  upon  the  flexing  of  a  diaphragm  or  bellows  as 
caused  by  varying  pressures  or  vacuums  to  obtain  the  mechanical  move- 
ment necessary  to  actuate  an  electrical  or  pneumatic  switch. 

Apparatus  Which  Operates  Valves 

Apparatus  which  is  so  mechanically  or  electrically  equipped  that  it  will 
open  and  close  valves,  and  possibly  give  them  fixed  intermediate  positions 
in  any  pipe  line  of  a  heating,  ventilating,  or  air  conditioning  system,  is 
termed  a  valve  operator.  The  function  of  a  valve  operator  is,  essentially, 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

to  move  the  plunger  of  a  valve  in  a  manner  required  by  its  type  and 
construction.  For  instance,  in  a  single-seat  valve,  the  disc  is  moved 
against  the  seat  and  held  there  with  sufficient  pressure  to  prevent  flow. 
A  three-way  valve,  however,  requires  a  valve  operator  that  will  place  the 
double  disc,  as  required,  between  the  two  seats.  Each  type  of  valve  has 
special  characteristics  to  which  a  valve  operator  must  be  adapted. 

When  a  valve  is  used  in  shut-off  service  the  valve  operator  simply 
opens  the  valve  or  closes  it  completely,  as  required.  When  the  valve  is  to 
provide  throttling  service,  a  different  type  of  valve  operator  is  used  so 
that  the  valve  may  be  held  at  any  intermediate  position  between  open 
and  closed.  Valve  operators  use  as  their  power  source  either  compressed 
air  (pneumatic  system),  electricity  (motor-driven  type  of  solenoid  type), 
or  a  volatile  liquid  (direct-connected  type). 

Apparatus  Which  Operates  Dampers 

Apparatus  which  is  so  mechanically  or  electrically  equipped  as  to  open 
and  close  dampers,  and  possibly  give  them  fixed  positions,  in  accordance 
with  the  purposes  of  the  system  using  the  dampers  is  termed  a  damper 
operator.  Damper  operators  are  made  for  opening,  closing,  and  position- 
ing the  dampers  in  the  ducts  of  heating,  ventilating,  or  air  conditioning 
systems  in  the  same  way  that  valve  operators  regulate  the  valves.  They 
receive  their  signals  from  thermostatic  or  manual  switches. 

The  sources  of  power  used  are  compressed  air,  electricity,  or  volatile 
liquids.  The  damper  operator  is  connected  to  its  damper  by  direct  con- 
nection or  by  a  linkage,  according  to  conditions,  and  it  can  usually  be 
mounted  either  outside  or  inside  the  duct  in  which  the  damper  is  located. 

Accessory  Apparatus 

Accessory  apparatus  is  that  additional  equipment  at  the  terminals  of  a 
control  system  necessary  to  make  it  operative.  Every  temperature  con- 
trol system  requires  a  number  of  accessories,  which  will  vary  with  the 
different  types  of  systems.  For  instance,  pneumatic  systems  require  a 
compressor  and  a  storage  tank  for  the  air  which  operates  the  units,  and 
low- voltage  electric  systems  require  a  .transformer  or  generator  to  provide 
the  required  current. 

Most  of  the  larger  control  systems  will  have  some  sort  of  central  switch- 
board which  may  include  indicating  and  recording  devices  as  well  as 
control  switches.  Thermostat  guards  are  generally  used  In  gymnasiums, 
schools,  and  places  of  assemblage  for  protective  purposes.  Time  switches 
and  similar  devices  are  often  important  parts  of  certain  types  of  control 
systems.  Couplings,  mountings,  and  indicators  are  often  parts  of  a 
system. 

Connecting  Apparatus 

Connecting  apparatus  is  that  equipment  used  to  connect  the  various 
parts  of  a  control  system.  Because  the  parts  of  the  system  are  often  some 
distance  apart,  the  connecting  means  are  important,  and  the  connections 
must  be  properly  planned  and  made. 

The  connecting  elements  are  fairly  obvious.  The  pneumatic  system 
uses  compressed  air  carried  in  small  pipes  and  tubing.  Electric  systems 

242 


CHAPTER  14 — AUTOMATIC  CONTROL 


are  wired  for  low-voltage  or  high-voltage  power  supply.  Systems  em- 
ploying volatile  liquids  generally  use  flexible  tubing  if  there  is  distance 
between  the  sensitive  ,bulb  and  the  operating  unit.  Each  form  has  certain 
limitations  which  the  designer  of  the  system  must  consider. 

Since  few  control  installations  are  alike,  the  manufacturers  of  control 
apparatus  usually  maintain  engineering  departments  staffed  by  experi- 
enced men  whose  advice  may  be  had  on  control  problems.  Progress  in 
automatic  control  has  been  rapid  in  the  past  few  years  and  the  field  of 
automatic  control  has  become  specialized. 

TEMPERATURE  CONTROL  SYSTEMS 

The  control  of  direct  radiation  is  simple.  Each  radiator  has  a  valve  on 
its  steam  or  water  supply,  with  a  thermostat  to  govern  the  opening  and 
closing  of  the  valve  to  maintain  the  desired  uniform  temperature.  One 
thermostat  may  control  the  valves  on  all  the  radiators  in  a  room,  or,  if  the 
room  is  large,  more  than  one  thermostat  may  be  used,  with  each  one 
governing  one  radiator  or  a  group  of  them.  Unit  type  thermostatic 
valves  may  be  used,  one  on  each  radiator. 

The  location  of  wall  thermostats  is  important.  They  must  be  on  inside 
walls  where  they  will  not  be  affected  by  drafts  of  either  warm  or  cold  air, 
but  where  they  will  be  exposed  to  general  room  conditions.  If  vibration 
is  present,  they  must  be  mounted  on  shock-absorbing  bases.  If  the  walls 
are  abnormally  hot  or  cold,  the  thermostats  must  be  mounted  on  heat- 
insulating  bases.  The  connecting  means  can  be  concealed  in  the  wall, 
under  the  floor  or  ceiling,  or  behind  baseboards  or  moldings. 

Modulating  type  valves  cannot  be  used  successfully  on  one-pipe  steam 
systems  because  the  partial  opening  of  valves  will  not  allow  the  con- 
densate  to  escape  against  the  incoming  steam. 

A  discussion  of  steam  heating  systems  is  given  in  Chapter  31,  and 
further  information  on  control  requirements  of  direct  radiation  may  be 
obtained  therefrom. 

Control  of  Unit  Heaters 

Unit  heaters  are  commonly  ceiling-hung  or  floor-mounted  units  con- 
sisting of  a  steam  or  hot  water  coil  with  a  fan  behind  it  to  force  air  past 
the  coil  and  into  the  room.  Vanes  direct  the  warm  air  flow.  The  simplest 
and  commonest  way  to  control  a  unit  heater  is  to  have  in  the  heated  space 
a  thermostat  which  will  turn  on  the  fan  when  heat  is  required  and  shut  it 
off  when  the  demand  is  satisfied.  However,  where  there  is  natural 
circulation  through  the  unit,  it  is  advisable  to  put  a  valve  on  the  steam  or 
hot  water  supply  line  and  arrange  it  so  the  steam  will  be  turned  on  only 
when  the  fan  is  running, 

As  a  precaution  against  allowing  the  unit  heater  motors  to  continue  to 
run  if  the  steam  supply  fails  or  is  for  some  reason  shut  off,  either  a  pres- 
surestat  or  a  thermostat  in  the  supply  line,  or  a  thermostat  on  the  return 
line  may  be  installed  to  stop  the  motor  when  the  pressure  or  temperature 
in  the  supply  line,  or  the  temperature  in  the  return  line,  drops  below  a 
predetermined  point.  When  the  fan  and  the  steam  are  controlled  simul- 
taneously, such  thermostat  will  also  prevent  the  blowing  of  cold  drafts. 

243 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

The  net  result  in  any  case  will  be  that  the  fan  will  run  only  when  there 
is  heat  in  the  coil. 

Control  of  Unit  Ventilators 

The  unit  ventilator  presents  a  different  control  problem  than  the  unit 
heater.  Generally  this  type  of  unit  draws  its  supply  of  air  from  the  out- 
side, heats  it,  and  introduces  this  air  into  the  room  under  control.  There 
are  many  types  of  unit  ventilators  on  the  market.  Some  have  a  mixing 
damper  by  which  the  temperature  of  the  air  entering  the  room  may  be 
varied,  others  have  valves  for  this  purpose,  and  still  others  use  a  com- 
bination of  the  two.  Regardless  of  the  construction  of  the  machine,  the 
essential  requirement  is  that  the  temperature  of  the  air  delivered  to  the 
room  should  change  slowly  and  remain  as  near  room  temperature  as 
possible.  Frequently  direct  radiators  are  used  in  conjunction  with  the 
unit  ventilators  to  supply  additional  heat  in  extremely  cold  weather  or 
for  quickly  heating  up  the  room. 

The  four  general  types  of  control  for  unit  ventilators  are  as  follows : 

1.  A  damper  operator,  which  is  controlled  by  a  room  thermostat,  is  attached  to  the 
mixing  damper.    When  the  thermostat  calls  for  heat,  the  damper  is  moved  to  a  position 
which  forces  more  air  through  the  heating  unit  and  thus  increases  the  amount  of  heat 
supplied  to  the  room.    This  action  must  be  gradual  so  that  the  air  temperature  may  be 
changed  slowly  to  prevent  the  drafty  condition  caused  by  supplying  first  hot  and  then 
cold  air.    This  simplest  arrangement  is  often  condemned  because  it  frequently  results 
in  drafts. 

2.  In  mild  weather  the  heating  unit  frequently  supplies  sufficient  heat  to  cause  over- 
heating of  the  room,  even  though  all  of  the  air  is  by-passed  around  the  heating  unit.    To 
avoid  this  fault  a  valve  is  placed  on  the  heating  unit  to  close  the  steam  supply  when  the 
damper  is  by-passing  all  of  the  air.    This  valve  is  used  in  addition  to  the  damper  operator 
explained  in  the  foregoing  paragraph,  but  though  giving  better  results,  it  may  fail  to 
prevent  drafts. 

3.  In  some  unit  ventilators  one  or  more  heating  units  are  used  without  a  mixing 
damper.   A  gradual-acting  valve  on  each  heating  unit  controls  the  supply  of  steam  to  the 
unit  to  give  the  proper  amount  of  heat  required  to  maintain  the  desired  room  tempera- 
ture.  A  thermostat  to  govern  each  valve  may  be  installed  in  the  room,  or  one^thermostat 
may  be  used  for  all  valves,  but  unless  a  thermostat  is  placed  directly  in  the  air  stream  of 
each  unit,  drafts  may  be  encountered. 

4.  Another  type  of  unit  ventilator  is  arranged  so  that  all  recirculated  air  passes 
through  the  heating  unit,  and  the  outside  air  is  introduced  into  the  room  for  cooling 
purposes  only.    The  outside  air  damper  and  the  recirculated  air  damper  are  interlocked 
so  that  one  damper  operator  will  control  them.    In  addition  a  valve  operator  is  placed 
on  the  heating  unit.    Both  of  the  operators  should  move  gradually  to  avoid  drafty  con- 
ditions.    When  the  thermostat  calls  for  heat,  the  damper  operator  slowly  closes  the 
outside  air  damper  and  simultaneously  opens  the  recirculating  damper;  if  this  does  not 
meet  the  demand,  the  valve  on  the  heating  unit  opens  until  the  room  temperature  reaches 
the  desired  point. 

For  additional  information  on  the  control  of  unit  ventilators,  refer  to 
Chapter  13. 

Central  Fan  Heating  and  Ventilating  Systems 

The  numerous  types  of  central  fan  systems  present  many  control 
problems.  In  general  they  all  have  one  point  in  common,  namely,  that 
the  temperature  change  may  be  very  fast  because  of  rapid  circulation. 

System  for  Ventilating  Only  (Split  System).  Fig.  4  shows  an  accepted 
control  for  ventilating  systems.  Thermostat  A  located  in  the  outside  air 

244 


CHAPTER  14 — AUTOMATIC  CONTROL 


duct  is  set  just  above  freezing,  and  controls  a  valve  C  on  the  first  heating 
coil.  This  valve  is  either  completely  open  or  completely  closed.  The  by- 
pass damper  B  and  the  other  two  valves  D  and  E  are  controlled  by  a  duct 
thermostat  F  located  in  the  discharge  duct  from  the  fan.  If  the  tempera- 
ture of  the  air  surrounding  the  thermostat  F  increases,  the  damper  is 
moved  automatically  to  admit  more  cold  air.  Should  this  not  reduce  the 
temperature  sufficiently,  the  valves  on  the  heating  coil  will  be  closed 
gradually  and  in  sequence  until  the  correct  temperature  is  reached.  The 
opening  or  closing  of  the  damper  B  and  the  valves  D  and  E  must  be 
gradual  or  there  will  be  a  wide  fluctuation  in  air  temperature. 

In  ventilating  systems  it  is  customary  to  supply  air  to  the  ventilated 
spaces  at  an  inlet  temperature  approximately  equal  to  the  temperature 
maintained  in  the  rooms.  The  radiators  therefore  are  designed  to  take 
care  of  all  the  heat  losses  from  the  room.  Hence,  in  order  to  maintain 


Electric  or  pneumatic 
power  source 


FIG.  4.    CONTROL  OF  A  SPLIT  SYSTEM  OF  VENTILATION 


controlled  room  temperatures  it  is  necessary  to  use  room  thermostats 
governing  control  valves  placed  on  the  radiators.  With  this  type  of 
central  fan  system  it  is  possible  to  ventilate  a  large  number  of  rooms  by 
means  of  one  fan. 

In  some  installations,  such  as  in  theaters  or  auditoriums,  it  is  difficult 
to  install  sufficient  direct  heating  surface  to  offset  the  heat  losses  from 
the  room.  Also  there  are  installations  where  a  short  heating-up  period  is 
allowed  before  occupancy  of  the  room,  and  it  is  advisable  to  use  the 
entire  heating  capacity  of  the  ventilating  system  for  this  purpose. 

In  central  fan  systems,  air  washers  are  often  used  and  in  such  cases,  due 
to  the  effect  of  temperatures  on  humidity,  additional  control  is  required. 
Fig.  5  shows  such  an  arrangement  with  control  of  the  second  tempering 
heating  unit  by  the  air  washer  temperature  and  with  the  usual  control 
of  the  first  tempering  heating  unit  by  the  outside  temperature.  This 
permits  the  air  to  be  kept  cool  while  passing  through  the  washer  so  that 
too  much  moisture  will  not  be  absorbed.  Fig.  5  also  shows  control  of  the 
reheating  units  by  a  duct  thermostat  in  the  fan  discharge,  and  the 
application  of  a  pilot  thermostat  to  a  system  of  this  sort. 

245 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

Combined  Systems.  There  are  various  central  fan  systems  which  are 
used  for  both  heating  and  ventilating.  They  are  usually  arranged  with 
tempering  heating  units,  automatically  controlled  to  provide  a  minimum 
temperature  for  ventilating  only,  and  additional  heating  units  to  supply 
the  heating  requirements.  Fig.  6  shows  a  type  of  system  which  has  the 
reheating  units  located  in  the  fan  room.  Tempered  air  at  about  70  F  is 
supplied  to  the  fan.  It  may  be  further  heated  by  the  reheating  units,  or  it 
may  pass  into  the  tempered  air  chamber.  A  room  thermostat  controls  a 
gradual-acting  damper  operator  on  the  double  mixing  damper  in  the  warm 
and  tempered  air  chambers.  When  the  thermostat  calls  for  heat,  the 


Electric  or  pneumatic 
power  source  / 


Multiple  point 
insertion  thermostat 

FIG.  5.    USE  OF  PILOT  THERMOSTAT  ON  VENTILATING  SYSTEM  WITH  AIR  WASHER 


damper  operator  moves  the  dampers  so  that  more  air  is  taken  from  the 
warm  air  chamber.  It  is  essential  that  the  double  mixing  damper  be 
moved  slowly  to  prevent  alternate  blasts  of  hot  and  cold  air  from  being 
supplied  to  the  room. 

Outside  Air,  Recirculating,  and  Vent  Dampers.  In  all  types  of  plenum 
systems,  the  outside  air  damper  is  usually  opened  and  closed  by  a  damper 
operator.  This  operator  may  be  controlled  from  a  switch  in  the  engi- 
neer's room  or  it  may  be  operated  by  a  relay  in  the  fan  motor  circuit. 
When  the  ventilating  fan  is  started,  the  relay  causes  the  damper  operator 
to  open  the  outside  air  damper.  ' 

Recirculating  dampers  and  vent  dampers  may  also  be  opened  and 
closed  by  means  of  damper  operators  controlled  from  remote  locations. 
Generally  these  damper  operators  are  positive  acting  and  are  either 
completely  opened  or  closed.  However,  in  some  cases  where  part  out- 
side air  and  part  recirculated  air  is  used,  it  is  advantageous  to  use  damper 
operators  which  have  a  certain  number  of  definite  positions.  With  this 
type  of  operator  it  would  be  possible  to  use  75  per  cent  outside  air  and 
25  per  cent  recirculated  air,  or  any  other  proportions  which  might  be 
predetermined.  These  damper  operators  are  controlled  from  switches 
generally  mechanically  interlocked  so  that  the  total  opening  of  the  two 
dampers  is  100  per  cent. 

246 


CHAPTER  14 — AUTOMATIC  CONTROL 


Hand- Fired  Coal  Systems 

In  small  buildings  the  heating  plant  may  be  controlled  by  a  single 
thermostat  located  in  a  key  room  in  the  building,  instead  of  each  room 
having  its  own  control. 

The  most  common  control  for  a  hand-fired  furnace  or  boiler  consists  of 
a  room  thermostat  and  a  furnace  regulator  of  some  type.  The  thermostat 
should  be  located  in  a  representative  room;  never,  of  course,  near  the 
chimney  or  heat  flue,  too  close  to  a  radiator,  or  in  a  drafty  hallway,  and 
preferably  on  an  inside  wall.  The  regulator  is  attached  to  the  draft  and 
check  dampers  of  the  furnace.  When  the  temperature  of  the  air  sur- 


FIG.  6.   CONTROL  OF  MIXING  DAMPERS  WITH  INTERMEDIATE-ACTING  THERMOSTAT 


rounding  the  thermostat  drops,  the  thermostat  causes  the  furnace  regu- 
lator to  open  the  draft  and  close  the  check  damper.  As  soon  as  the  room 
comes  up  to  temperature,  the  draft  is  closed  and  the  check  damper 
opened.  With  this  arrangement  on  hot  water  heating  systems  it  is 
advisable  to  install  an  immersion  thermostat  in  the  boiler.  This  thermo- 
stat should  be  connected  with  the  room  thermostat  so  that  both  must  call 
for  heat  before  the  draft  is  opened,  but  either  one  may  cause  the  draft  to 
be  closed.  On- warm  air  systems  it  is  advisable  to  use  a  bonnet  thermostat 
and  on  steam  heating  systems  a  pressure  limiting  device,  in  series  in  each 
case  with  the  room  thermostat.  If  the  temperature  of  the  heating 
medium  becomes  too  high,  the  drafts  will  be  closed  even  though  the  room 
thermostat  continues  to  call  for  heat* 

There  have  been  some  recent  improvements  in  controls  of  this  type, 
involving  the  use  of  special  types  of  thermostats  and  auxiliary  apparatus 
which  will  give  closer  control  and  prevent  overheating  in  mild  weather. 

CONTROL  OF  AUTOMATIC  FUEL  APPLIANCES 

It  is  essential  that  automatic  temperature  control  be  used  with  oil 
burners,  gas  burners,  and  stokers  to  aid  economical  operation.  There  are 

247 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

many  types  of  burners  and  many  types  of  control,  but  there  are  some 
points  common  to  all.  First,  a  room  thermostat  is  located  in  a  key 
position  in  the  building  to  maintain  a  given  temperature  at  that  point. 
Safety  devices  are  installed  in  connection  with  this  thermostat  so  that  a 
failure  of  the  ignition,  power,  or  fuel  supply  will  shut  the  system  down. 
The  same  limit  controls  as  recommended  for  coal  burning  should  be 
used. 

Oil  Burners 

Fig.  7  illustrates  diagrammatically  the  essentials  of  an  oil  burner  con- 
trol circuit.  Three  thermostats  are  employed  as  shown  in  the  illustration. 
Thermostat  No.  1  will  stop  the  burner  when  the  room  temperature  is  too 
high  and  No.  2  will  stop  the  burner  when  the  temperature  of  the  heating 
medium  exceeds  the  setting  of  thermostat  No.  2.  Both  temperatures 
must  be  below  their  respective  thermostat  settings  to  start  the  burner. 
Thermostat  No.  3  responds  to  the  flame  temperatures  and  in  conjunction 
with  the  control  switch  acts  as  a  safety  to  stop  the  burner  if  the  latter 
fails  to  ignite  or  burn  properly  as  demanded  by  thermostats  No.  1  and  2. 

Domestic  Applications 

Steam  and  hot  water  heating  plants  are  often  used  to  provide  heat  for 
the  domestic  hot  water  supply  as  well  as  for  heating  the  building.  Fig  8 
illustrates  one  such  system.  The  burner  control  is  similar  to  that  shown 
in  Fig.  7  except  that  either  the  room  thermostat  or  the  tank  thermostat 
may  start  the  burner.  If  the  house  is  warm  enough,  the  house  tempera- 
ture control  valve  will  remain  closed,  and  the  boiler,  through  the  coil 
heater,  will  warm  the  water  in  the  storage  tank  when  the  tank  thermostat 
starts  the  burner.  The  burner  will  stop  only  when  both  thermostats  are 
satisfied,  or  when  the  steam  pressure  shall  have  reached  that  allowed  by 
the  pressurestat.  Much  the  same  control  is  applied  to  gas  burners  and 
automatic  coal  stokers. 

Gas  Heating  Appliances 

On  account  of  the  ease  and  effectiveness  with  which  the  fuel  can  be 
controlled,  gas-burning  appliances  are  particularly  adaptable  to  full 
automatic  control.  Standard  equipment  on  a  steam  boiler  generally  in- 
cludes provision  for  control  through  a  room  temperature  thermostat,  a 
steam  pressure  regulator,  and  a  device  which  shuts  off  the  gas  in  the  event 
that  the  water  level  becomes  too  low.  Practically  all  gas  boilers  are  or 
may  be  equipped  with  automatic  safety  pilots  which  shut  off  the  gas  if  the 
pilot  flame  is  too  low. 

Water  boilers  are  adapted  to  operation  under  thermostatic  room  tem- 
perature control  and  are  also  provided  with  water  temperature  control 
equipment.  Warm  air  furnaces  can  be  under  the  control  of  thermostats 
in  the  spaces  being  heated,  as  well  as  thermostats  located  in  the  heat  ducts 
for  the  purpose  of  preventing  unpleasantly  hot  air  reaching  the  heated 
spaces.  Variations  in  the  pressure  under  which  the  gas  is  supplied  to  the 
appliance  are  controlled  by  means  of  a  gas-pressure  regulator.  This  is  an 
essential  part  of  practically  all  makes  of  gas-burning  heating  appliances ; 
in  fact,  a  gas-pressure  regulator  is  required  by  the  American  Gas  Associa- 
tion on  all  approved  gas  boilers,  warm  air  furnaces  (except  floor  furnaces) , 
and  unit  heaters. 

248 


CHAPTER  14 — AUTOMATIC  CONTROL 


INDIVIDUAL  ROOM  CONTROL 

The  most  elaborate  type  of  automatic  control  is  that  by  which  the 
temperature  in  each  room  or  in  a  group  of  rooms  can  be  controlled.  A 
thermostat  in  each  room  governs  the  valves  on  the  radiators  in  that  room, 


FIG.  7.    ELECTRIC  THERMOSTAT  APPLIED  TO  OIL-FIRED  HEATING  SYSTEM 


Room 
Thermostat 


To  Hot  Water 


lank 

^Thermostat  Pressure-^ 
,  <5tat      \ 


House  Temperature 
"Control  Vafve 


^  To  Radiation 


-Mbfart '=^,rrom  Radiation 

L/rre 


^AutofT7at/c  Fuel  Burner 


FIG.  8.   TYPICAL  ARRANGEMENT  OF  STEAM  OR  VAPOR  SYSTEM  WITH  Two 

THERMOSTATS  CONTROLLING  AUTOMATIC  FUEL  BURNER  USED 

FOR  HOUSE  HEATING  AND  WATER  HEATING 

opening  them  as  heat  is  called  for  and  shutting  them  when  the  room  is 
warm  enough.  The  thermostats  are  all  connected  in  relay  so  when  any 
thermostat  is  calling  for  heat,  an  automatic  burner  will  supply  steam,  hot 
water,  or  warm  air,  to  the  system ;  and  when  all  the  thermostats  are  satis- 
fied, the  burner  will  shut  off.  This  is  an  excellent  arrangement  for  larger 

249 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

residences,  and  it  may  be  applied,  in  modified  form,  in  houses  which  have 
one  room  or  a  section  that  is  difficult  to  heat. 

ZONE  CONTROL 

Zone  control  is  a  step  between  a  single  thermostat  and  individual  room 
temperature  control.  The  building  is  first  divided  into  sections  or  zones 
which  may  have  quite  different  heat  requirements.  With  this  method  of 
control : 

First:  The  zoning  should  be  done  with  reference  to  the  compass,  since 
the  north  and  west  quarters  in  most  localities  require  considerably 
more  heat  during  the  heating  season  than  do  the  south  and  east 
quarters. 

Second:  Most  large  office  buildings  have  more  or  less  space  occupied 
by  merchants,  and  some  by  clubs,  or  restaurants,  which  have  short 
hours  of  occupancy.  Much  can  be  accomplished  in  zoning  with 
reference  to  the  kind  of  occupancy  of  space.  For  additional  infor- 
mation on  this  subject,  refer  to  Chapter  31. 

Variations  of  the  usual  zone  control  methods  by  the  use  of  recently 
developed  special  devices  have  been  quite  successful  in  obtaining  greater 
economy  from  heating  systems.  Frequently  these  use  an  outside  ther- 
mostat or  group  of  thermostats  which  adjust  the  operation  of  the  controls 
to  conform  to  variations  in  weather  conditions. 

COOLING  UNITS 

Cooling  units  are  readily  adaptable  to  thermostatic  control.  Several 
arrangements  are  as  follows: 

1.  Room  thermostat  in  conjunction  with  a  magnetic  or  motor-operated  valve  to 
regulate  the  flow  of  refrigerant  to  coil.    Usually  the  fans  operate  continuously. 

2.  Room  thermostat  to  control  the  operation  of  the  compressor.    The  fans  operate 
continuously. 

3.  Room  thermostat  to  control  the  operation  of  the  fan  motors. 

4.  Room  thermostat  to  control  the  operation  of  the  fan  motor  and  the  compressor 
motor  simultaneously. 

5.  Room  thermostat  to  control  the  operation  of  the  compressor  with  back  pressure 
control  to  regulate  the  fans. 

INDUSTRIAL  PROCESSES 

There  are  many  industrial  processes  requiring  automatic  temperature 
and  humidity  regulation.  The  control  equipment  operates  on  the  same 
principles  that  have  been  described,  but  it  is  often  especially  designed  for 
each  particular  process.  Each  installation,  or  the  installation  for  each 
process,  is  likely  to  be  a  problem  peculiar  to  that  process. 

AIR  CONDITIONING  SYSTEMS 

The  following  fundamental  principles  should  be  borne  in  mind  in  the 
solution  of  problems  involving  the  control  of  air  conditioning  systems: 

250 


CHAPTER  14 — AUTOMATIC  CONTROL 


1.  Dew-point  temperatures  vary  only  with  the  amount  of  moisture.     That  is,  no 
matter  how  much  a  given  mixture  of  air  and  water  vapor  is  heated  or  cooled,  the  dew- 
point  temperature  remains  the  same,  as  long  as  there  is  no  addition  or  subtraction  of 
water.    Cooling  below  the  dew-point  temperature  will,  of  course,  cause  condensation  of 
the  water  vapor.    Also,  at  the  same  temperature,  there  is  always  the  same  proportion  of 
water  vapor  in  the  saturated  mixture,  provided  sufficient  water  and  time  are  furnished 
for  saturation. 

Table  5,  Chapter  1.  shows  the  amount  of  moisture  required  to  saturate  a  space  at 
various  temperatures.  When  the  proper  amount  of  moisture  is  determined,  it  is  only 
necessary  to  set  the  air  washer  (dew-point)  thermostat  for  the  corresponding  temperature 
of  saturation;  then  if  the  air  Centering  the  washer  has  more  humidity  than  desired,  the 
excess  will  be  condensed ;  and  if  it  has  less,  the  deficiency  will  be  absorbed  from  the  sprays. 

For  example,  the  dew-point  temperature  at  70  F  and  40  per  cent  relative  humidity  is 
45  F.  Therefore,  if  the  air  temperature  is  maintained  at  45  F  as  it  leaves  an  air  washer 
(assuming  it  is  fully  saturated)  and  then  is  heated  to  70  F,  it  will  have  a  relative  humidity 
of  40  per  cent.  If  it  is  desired  to  maintain  these  conditions  in  a  given  space,  the  air  tem- 
perature can  be  raised  to  any  necessary  point,  say  120  F  (at  which  the  relative  humidity 
will  be  only  9  per  cent) .  When  the  heat  in  the  air  has  been  dissipated,  the  space  tem- 
perature being  maintained  at  70  F,  the  relative  humidity  will  be  40  per  cent. 

2.  Within  ordinary  operating  ranges,  saturated  air  will  have  a  relative  humidity  of 
approximately  50  per  cent  when  its  temperature  is  raised  20  deg.    For  example,  satu- 
rated air  at  40  F  raised  to  60  F  has  a  relative  humidity  of  48  per  cent;  60  F  saturated  air 
raised  to  80  F  has  a  relative  humidity  of  50  per  cent.    (See  Table  4,  Chapter  1.)    Thus 
a  differential  thermostat  can  be  used  to  maintain  a  nearly  constant  relative  humidity  of 
50  per  cent  by  holding  the  dew-point  temperature  20  deg  below  the  dry-bulb  temperature. 

3.  The  total  heat  of  the  air  and  the  water  vapor  mixed  with  it  varies  directly  with  the 
wet-bulb  temperature.    For  example,  the  occupants  of  an  auditorium  give  off  sensible 
heat  which  tends  to  raise  both  the  dry-bulb  and  the  wet-bulb  temperatures  of  the  space ; 
but  the  occupants  also  give  off  moisture  which  increases  the  absolute  humidity  and  tends 
to  further  raise  the  wet-bulb  temperature  by  an  amount  which  is  a  direct  indication  of 
the  heat  expended  by  each  occupant  in  evaporating  this  water.    This  relationship  is 
useful  in  regulating  the  total  heat,  as  wet-bulb  temperatures  can  be  controlled  directly 
by  means  of  a  thermostat  having  a  sensitive  element  covered  with  water-fed  wicking, 
similar  to  a  wet-bulb  thermometer. 

For  example,  the  total  heat  of  air  at  80  F  and  60  per  cent  relative  humidity  is  the  same 
as  for  air  saturated  at  70  F,  i.e.,  33.5  Btu  per  pound,  both  having  a  wet-bulb  temperature 
of  70  F.  Air  at  80  F  and  60  per  cent  relative  humidity  (70  F  wet-bulb  =  33.5  Btu  per 
pound)  reduced  to  70  F  and  50  per  cent  relative  humidity  (58}^  F  wet-bulb  =  25.2  Btu 
per  pound,  total  heat)  must  give  up  8.3  Btu  per  pound.  If  the  sensible  heat  and  moisture 
pick-up  in  an  auditorium  is  8.3  Btu  per  pound  of  air  handled  in  the  conditioning  system, 
the  wet-bulb  temperature  of  the  air  entering  the  space  must  be  maintained  at  58J^  F  to 
secure  a  final  condition  of  80  F  and  60  per  cent  relative  humidity. 

Control  of  Relative  Humidity 

The  following  are  the  most  commonly  used  methods  of  controlling 
relative  humidity: 

1.  A  thermostat  is  located  in  or  at  the  outlet  of  a  spray-type  air  conditioner  which 
maintains  a  constant  saturation  temperature  of  the  air  leaving  the  conditioner  by  varying 
the  temperature  of  water  entering  the  suction  of  the  pump^  supplying  the  spray  nozzles, 
or  by  varying  the  temperature  of  the  air  entering  the  conditioner,  or  both.   The  tempera- 
ture of  the  air  entering  the  conditioner  may  be  varied  by  use  of  tempering  heaters,  or  by 
the  proper  proportioning  of  supply  and  return  air  entering  the  conditioner.   This  thermo- 
stat is  known  as  a  dew-point  thermostat,  as  it  determines  the  dew-point  temperature  of 
the  air  introduced  into  the  conditioned  spaces.     A  second  thermostat  in  the  room,  or  in 
the  path  of  the  air  leaving  the  room,  maintains  a  constant  dry-bulb  temperature  by 
varying  the  amount  of  sensible  heat  added  to  the  air  leaving  the  conditioner,  or  by 
varying  the  volume  of  air  introduced  into  the  conditioned  spaces.    These  two  ther- 
mostats, in  combination,  control  the  dry-bulb  and  dew-point  temperatures,  which 
accordingly  fix  the  relative  humidity. 

2.  A  wet-bulb  thermostat  is  located  in  the  room,  or  in  the  path  of  the  air  leaving  the 
roomf  to  maintain  a  constant  wet-bulb  temperature  by  varying  the  saturation  tempera- 

251 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

ture  at  the  air  conditioner  outlet.  A  dry-bulb  thermostat  is  located  in  the  room  to 
maintain  a  constant  dry-bulb  temperature,  which  in  combination  with  a  constant  wet- 
bulb  temperature  fixes  the  relative  humidity. 

3.  A  differential  thermostat  may  be  used  to  control  relative  humidity.    This  instru- 
ment consists  of  two  thermostatic  elements,  one  of  which  is  in  the  path  of  the  air  leaving 
the  conditioner,  and  the  other  under  the  influence  of  the  dry-bulb  temperature  in  the 
room.    Instruments  of  this  kind  maintain  a  constant  relative  humidity  by  maintaining 
a  constant  difference  between  the  dew-point  temperature  and  the  dry-bulb  temperature 
in  the  room*    (See  Item  2  under  Air  Conditioning  Systems.)    One  thermostatic  element 
may  be  equipped  with  a  moistening  device  to  permit  it  to  operate  on  wet-bulb  tem- 
peratures.   Such  an  instrument  can  be  used  to  control  the  wet-bulb  depression  and  thus 
the  relative  humidity. 

4.  A  humidistat  which  responds  directly  to  changes  in  humidity  may  be  used  to 
maintain  a  predetermined  relative  humidity  with  constant  or  with  varying  temperature. 
It  may  do  this  by  varying  the  dew-point  temperature  of  air  leaving  a  conditioner;  by 
varying,  with  dampers,  the  proportion  of  moist  and  dry  air;  by  varying  the  amount  of 
moisture  otherwise  added  to  the  air;  or  by  varying  the  dry-bulb  temperature. 

Humidificarion  for  Residences 

The  principles  underlying  humidity  requirements  and  limitations  for 
residences  are  summarized  in  University  of  Illinois  Bulletin  No.  481,  as 
follows: 

1.  Optimum  comfort  is  the  most  tangible  criterion  for  determining  the  air  conditions 
within  a  residence. 

2.  An  effective  temperature  of  65  deg2  represents  the  optimum  comfort  for  the 
majority  of  people.    Under  the  conditions  in  the  average  residence  a  dry-bulb  tempera- 
ture of  69.5  F  with  relative  humidity  of  40  per  cent  is  the  most  practical  for  the  attain- 
ment of  65-deg  effective  temperature. 

3.  Evaporation  requirements  to  maintain  a  relative  humidity  of  40  per  cent  in  zero 
weather  depend  on  the  amount  of  air  inleakage  to  the  average  residence,  and  vary  from 
practically  nothing  to  24  gal  of  water  per  24  hours. 

4.  Relative  humidity  of  40  per  cent  indoors  cannot  be  maintained  in  rigorous  climates 
without  excessive  condensation  on  the  windows  unless  tight-fitting  storm  sash  or  the 
equivalent  is  installed. 

5.  The  problems  of  humidity  requirements  and  limitations  cannot  be  separated  from 
considerations  of  good  building  construction,  and  the  latter  should  receive  serious  atten- 
tion in  the  installation  of  humidifying  apparatus. 

The  following  conclusions  were  drawn  from  the  experimental  results 
reported  in  the  aforementioned  bulletin: 

1.  None  of  the  types  of  warm  air  furnace  water  pans  tested  proved  adequate  to 
evaporate  sufficient  water  to  maintain  40  per  cent  relative  humidity  in  the  Research 
Residence  except  only  in  moderately  cold  weather. 

2.  The  water  pans  used  in  the  radiator  shields  tested  did  not  prove  adequate  to  main- 
tain 40  per  cent  relative  humidity  in  a  residence  similar  to  the  Research  Residence  when 
the  outdoor  temperature  approximated  zero  degrees  Fahrenheit. 

Central  Fan  Air  Conditioning  Systems 

In  central  fan  air  conditioning  systems  as  described  in  Chapters  9  and 
22,  varying  amounts  of  outside  and  recirculated  air  are  used,  except  where 
contamination  prevents  re-use,  and  in  general  for  obtaining  humidity 
control  under  winter  conditions  heat  is  supplied  to  the  air  after  it  has 
passed  the  air  washer.  There  are  many  control  variations  in  use,  and 

1See  Humidification  for  Residences,  by  A,  P.  Kratz  (University  of  Illinois,  Bulletin  No.  48). 
^Sixty-six  deg  is  the  optimum  winter  effective  temperature  recommended  by  the  A.S.H.V.E.  Committee 
on  Ventilation  Standards.    See  Chapter  2. 

252 


CHAPTER  14 — AUTOMATIC  CONTROL 


Fig.  9  shows  a  composite  diagram,  rather  than  a  system  of  control  for  a 
single  installation.  The  control  valves  for  a  dehumidifying  air  washer  are 
shown  in  Fig.  10.  The  functions  of  the  control  devices  shown  in  Figs.  9 
and  10  are  as  follows: 

Winter  Operation  (With  Steam) 

1.  Thermostat  A  opens  a  direct-acting  valve  in  the  steam  supply  to  a  low-capacity 
tempering  coil  P.    The  thermostat  is  set  at  35  F. 

2.  Thermostat  B  in  the  path  of  the  air  leaving  the  second  tempering  coil  Q  controls  a 
valve  in  the  steam  supply  to  the  coil  Q  at  45  F. 

3.  Thermostat  C  controls  the  intake  M  and  return  air  N  dampers  at  50  F.    This 
location  of  thermostat  C  is  primarily  for  operation  with  steam  heating  and  at  such  times 
as  by-pass  damper  0  is  closed.    See  discussion  under  the  heading  Spring  and  Fall  Opera- 
tion . 


FIG.  9. 


DIAGRAMMATIC  ARRANGEMENT  OF  VARIOUS  PHASES  OF  CONTROL  FOR  A 
CENTRAL  FAN  AIR  CONDITIONING  SYSTEM 


4.  Humidistat  or  wet-bulb  thermostat  D  in  the  return  air,  acting  through  a  relay, 
causes  C  to  call  for  outside  air  when  the  relative  humidity  rises  above  55  per  cent  or  the 
wet-bulb  temperature  rises  above  60  F;  also,  if  necessary,  thermostat  D  shuts  off  the 
water  supply  to  the  spray  heads  in  the  air  washer  and  opens  the  supply  to  the  flooding 
nozzles  at  the  eliminator  plates,  by  operating  the  three-way  valve  U  (Fig.  10).    The 
relative  humidity  must,  of  course,  be  changed  to  suit  the  requirements.    It  must  be 
maintained  low  enough  to  avoid  condensation  on  walls  or  windows.3 

5.  Thermostat  E  in  the  discharge  end  of  the  air  washer  operates  a  three-way  valve 
( V,  Fig.  10)  in  the  water  circulating  line  so  as  to  cause  water  to  pass  through  or  around 
a  heating  unit  in  order  to  produce  the  correct  dew-point  temperature  by  adding  any 
necessary  heat  to  the  water.   It  may  also  operate  a  reverse  valve  W  (Fig.  10)  in  the  steam 
supply  to  the  heating  unit.    The  heat  added  may  be  only  that  sufficient  to  make  up  the 
temperature  drop  through  the  washer  caused  by  evaporation.     This  thermostat  is 
reverse-acting  to  prevent  over-humidification  in  case  of  failure  of  the  motive  power. 

6.  Thermostat  F  in  the  fan  discharge  operates  a  valve  in  the  steam  supply  to  the 
heater  R  in  order  to  produce  the  lowest  temperature  at  which  air  can  be  introduced  into 
the  conditioned  space,  without  complaints  of  draft.    This  varies  from  60  F  to  70  F, 
depending  on  the  velocity  through  the  supply  grilles  and  their  location. 


"See  discussion  of  condensation  in  Chapter  7.      Also  see  paper  entitled  Frost  and  Condensation  on 
Windows,  by  L.  W.  Leonhard  and  J.  A.  Grant  (A.S.H.V.E.  TRANSACTIONS,  Vol.  35, 1929). 

253 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

7.  Room  thermostat  G  in  a  representative  location  controls  a  valve  in  the  steam  supply- 
to  the  coil  or  coils  S  which  supply  the  heat  to  replace  the  loss  from  the  conditioned  space. 

Summer  Operation  (With  Refrigeration) 

Thermostats  A,  B,  F,  and  G  all  hold  their  valves  closed  during  summer 
temperatures  which  are  above  the  thermostat  settings,  although  this  is 
unimportant  while  no  steam  is  being  supplied. 

1.  Thermostat  C,  having  been  set  for  50  F,  supplies  power  to  open  wide  the  intake 
damper  and  close  the  return  air  damper  under  the  higher  summer  temperatures,  and  this 
power  can  be  passed  through  a  graduating  switch  to  permit  manual  operation  of  the 
dampers.    As  the  wet-bulb  temperature,  or  total  heat,  of  the  outdoor  air  is  now  normally 
greater  than  that  of  the  return  air,  it  is  desirable  in  order  to  keep  down  cooling  costs 
to  recirculate  the  maximum  amount  of  air. 

2.  Humidistat  D  is  by-passed  so  that  power  is  applied  directly  to  the  three-way  valve 
U  (Fig.  10)  to  prevent  shutting  off  the  sprays.    This  by-pass  can  be  arranged  for  cutting 
in  manually,  or  automatically,  with  the  starting  of  the  refrigerating  machinery. 

t-To  Sprays  To  Flooding  Nozzles 

I 
Three-way  Vatve  U-+.\ 


Three-way  VaJve  V>*.Q 
Reverse  Valve  W+ 
Air  Flow  -4*1 


_       -? 
I 


Cooling  Ta-nkT 


Heater 


Pump- 
FIG.  10.    CONTROL  VALVES  FOR  A  DEHUMIDIFYING  AIR  WASHER 

3.  Thermostat  E,  operating  the  three-way  valve  V  (Fig.  10),  now  determines  whether 
the  spray  water  is  to  be  passed  through  refrigerated  coils  or  is  to  be  recirculated  without 
treatment,  and  thus  it  regulates  the  dew-point  temperature.    It  is  assumed  that  steam 
and  refrigeration  are  not  both  turned  on  at  the  same  time. 

4.  Thermostat  H  operates  a  damper  0  in  the  by-pass  space  around  the  air  washer  so 
as  to  mix  the  warmer  return  air  with  the  cold  air  leaving  the  dehumidifier  in  such  pro- 
portions as  to  give  the  minimum  temperature  at  which  air  can  be  introduced  into  ^the 
conditioned  space.     This  might  be  70  F  for  a  room  temperature  of  85  F.     A  switch 
installed  in  the  power  line  from  H  should  be  so  connected  as  to  permit  keeping  damper  0 
closed  during  winter  operation. 

5.  Thermostat  G,  in  addition  to  operating  a  valve  on  the  heating  unit  S,  acts  as  a  pilot 
for  thermostat  H  so  as  to  retard  the  action  of  the  latter  in  closing  the  by-pass  damper 
when  the  temperature  in  the  space  is  below  the  desired  point. 

Spring  and  Fall  Operation 

During  a  considerable  part  of  the  year,  conditioning  can  be  ^accom- 
plished by  using  all  outside  air  or  by  mixing  it  with  returned  air.  For 
example,  when  the  total  sensible  heat  gain  in  an  auditorium  is  2.4  Btu  per 
pound  of  air  being  treated,  outside  air  will  be  raised  from  60  F  to  70  F  by 
the  heat  gain.  During  this  period  when  dry-bulb  temperatures  are^to  be 
maintained  at,  or  not  much  above,  70  F,  the  gain  in  sensible  heat  is  the 
only  factor  that  need  be  considered,  because  it  is  large  in  comparison  with 
the  gain  in  latent  heat,  except  in  restaurants  and  in  some  classes  of 
industrial  work.  The  intake  and  recirculating  dampers  can  then  be 
operated  by  thermostat  F  set  at  60  F,  It  is  assumed  that  such  an  .outlet 

254 


CHAPTER  14 — AUTOMATIC  CONTROL 


temperature  can  be  used;  if  not,  the  volume  of  air  should  be  increased. 
Thermostat  H,  being  set  higher  for  hot  weather,  holds  by-pass  damper  0 
open  to  provide  a  maximum  volume  of  air.  In  order  to  minimize  over- 
humidifl cation,  the  air  washer  and  by-pass  are  arranged  so  that  the  return 
air  stream  tends  to  use  the  by-pass.  However,  since  dehumidification  is 
not  required,  the  humidity  control  is  obtained  by  shutting  off  the  spray 
water  by  humidistat  D. 

Except  for  heating-up  periods  or  other  times  when  the  heat  gain  is  not 
greater  than  the  heat  loss,  a  system  of  this  type  can  be  operated  without 
artificial  heat  with  outdoor  temperatures  as  low  as  40  F.  For  this  reason 
it  is  economical  to  place  a  thermostat  in  the  return  air  near  D  set  to  shut 
off  a  valve  in  the  main  steam  supply  to  the  system  at  a  temperature 
about  3  degrees  below  that  desired  in  the  conditioned  space.  A  pilot 
thermostat  exposed  to  the  outdoor  temperature  prevents  the  shut-off  on 
days  colder  than  40  F. 

As  previously  stated,  there  can  be  many  variations  from  these  descrip- 
tions, some  of  which  are : 

1.  Tempering  coils  may  consist  of  only  one  bank,  P  or  Q,  controlled  by  thermostat  A 
or  thermostat  B.  In  any  case  the  capacity  of  the  heating  unit  controlled  by  the  outdoor 
temperature  must  be  as  low  as  feasible,  otherwise  if  steam  is  supplied  to  it  when  the  out- 
door temperature  is  30  F,  the  temperature  of  the  air  entering  the  washer  is  likely  to  be  too 
high  to  permit  maintaining  the  proper  dew-point  temperature. 

_  2.  Both  tempering  coils  may  be  omitted  and  the  return  air  may  be  mixed  with  outside 
air  by  thermostat  C  so  as  to  provide  a  proper  temperature  at  the  washer  inlet.  In  this 
case,  humidistat  D  should  not  act  as  a  pilot. 

3.  The  heating  unit  for  the  air  washer  water  may  be  omitted,  and  the  proper  dew-point 
temperature  maintained  by  placing  thermostat  C  in  the  location  of  E.    This  requires 
either  additional  heat  from  the  tempering  coils  or  more  return  air  to  make  up  the  loss  due 
to  evaporation  in  the  washer. 

4.  Heating  unit  5  may  be  combined  with  R  in  one  or  two  banks  and  controlled  by  a 
one-  or  two-point  thermostat  at  F,  set  for  the  minimum  temperature  at  which  air  can  be 
admitted  into  the  conditioned  space.    For  heating  purposes,  thermostat  G  then  becomes 
a  pilot  for  F  so  that  these  heating  units  are  operating  at  full  capacity  when  the  space  is 
cold,  and  are  throttled  by  F  when  no  heat  is  required. 

5.  Another  arrangement  is  the  use  of  a  type  of  thermostat  at  F  which  can  operate 
at   any  temperature  between  a  proper  minimum  and  a  necessary  maximum,    de- 
pending on  the  temperature  of  the  space.    Thus  for  winter  operation  when  the  room 
temperature  is  68  F,  the  blower  delivers  air  sufficiently  warm  to  supply  the  heat  required 
under  extreme  conditions,  and  when  it  is  74  F,  the  delivery  will  be  as  cool  as  possible 
without  complaint  of  drafts.    A  similar  device  can  be  used  to  replace  H,  and  be  set  to 
operate  between  60  and  80  F  for  summer  conditions. 

6.  For  summer  use,  a  remote  readjustable  thermostat  can  be  located  at  H,  and  can  be 
reset  by  a  pilot  exposed  to  the  outdoor  temperature.    Thus  as  the  outdoor  temperature 
increases,  the  temperature  in  the  space  is  maintained  at  a  higher  point. 

7.  A  constant  portion  of  the  return  air  may  be  brought  to  a  point  between  the  air 
washer  and  the  blower,  and  the  temperature  of  the  air  leaving  the  washer  may  be  regu- 
lated to  give  the  proper  result  at  H.    The  regulation  is  accomplished  by  shutting  off  one 
or  more  groups  of  sprays,  or  by  changing  the  temperature  of  the  spray  water  until  the 
proper  degree  of  cooling  is  secured. 

8.  Where  an  air  washer  large  enough  to  pass  all  the  air  handled  by  the  fan  is  selected, 
the  by-pass  and  its  damper  0  are  not  used.    The  washer  sprays  must  be  divided  into  two 
side-by-side  sections  so  that  one  section  can  be  turned  on  or  off  by  H  to  provide  the 
proper  temperature. 

9.  Where  an  ejector  type  heating  unit  is  used  for  the  spray  water,  a  reverse-acting 
valve  similar  to  W  (Fig.  10)  must  be  placed  in  the  steam  supply  to  be  operated  by  ther- 
mostat E.    In  this  case  it  is  usual  to  install  in  this  steam  line  another  reverse-acting 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

valve  to  be  operated  directly  by  the  water  pressure  in  the  pump  discharge  line.    This 
automatically  shuts  off  the  steam  when  the  water  circulating  pump  is  not  in  operation. 

10.  Based  on  the  fact  that  the  spray  water  in  the  air  washer  pan  has  practically  the 
same  temperature  as  the  air  leaving  the  washer,  dew-point  control  can  be  accomplished 
by  installing  thermostat  E  in  the  water  pan. 

11.  Where  cold  well-water  is  used  for  dehumidification,  it  is  admitted  to  the  sprays 
through  a  three-way  valve  similar  to  V  which  is  operated  by  thermostat  E. 

12.  Control  of  steam  heat  is  shown  entirely  by  valves,  although  it  is  usual  to  install  a 
by-pass  damper  around  each  heating  unit  and  to  operate  it,  either  with  or  without  a 
damper  over  the  face  of  the  heating  unit,  in  conjunction  with  the  valve. 


PROBLEMS  IX  PRACTICE 

1  •  How  may  temperature  control  be  obtained  in  a  room  heated  by  a  radiator 
with  a  constant  steam  supply? 

By  a  thermostat  handling  an  individual  radiator  valve  pneumatically  or  electrically,  or 
by  a  self-contained  radiator  valve. 

2  •  How  may  temperature  control  be  obtained  in  a  room  heated  by  a  unit 
heater? 

With  constant  steam  supply,  the  unit  heater  motor  may  be  started  or  stopped  by  a 
thermostat,  either  directly  or  through  a  relay.  With  intermittent  steam  supply,  opera- 
tion of  the  motor  by  thermostat  can  be  limited  to  the  time  that  steam  is  available,  by 
using  a  reverse-acting  temperature  or  pressure  limit  switch. 

3  •  How  may  temperature  control  be  provided  in  a  room  heated  and  venti- 
lated by  a  unit  ventilator  which  includes  two  ex  tended -surf ace  units? 

Operation  of  the  unit  for  service  during  occupancy  of  the  room  may  be  manual,  by 
switch,  or  by  time  clock.  When  the  desired  temperature  level  is  reached,  the  outside 
air  intake  may  be  controlled  by  a  damper  motor  coupled  with  the  fan  motor  circuit  by 
means  of  a  thermostat.  The  outside  air  damper  will  operate  to  a  given  position  in 
either  case. 

Air  passing  through  the  unit  may  be  preheated  through  the  first  heating  coil  to  a  definite 
temperature  by  a  control  valve  on  the  steam  supply  governed  by  a  temperature  controller 
reacting  to  the  temperature  of  the  air  on  the  outlet  side  of  the  convector.  The  second 
heating  coil  may  provide  the  necessary  heating  capacity,  and  the  steam  supply  to  this 
coil  may  be  modulated,  either  manually  or  automatically,  in  accordance  with  the  tem- 
perature required  in  the  room. 

4  •  How  may  temperature  control  be  obtained  in  a  room  heated  by  a  duct 
system? 

Air  may  enter  the  room  from  the  central  fan  system  at  a  predetermined  minimum  tem- 
perature. Heaters  placed  in  the  duct  to  bring  the  air  up  to  this  temperature  should  be 
equipped  with  face  and  by-pass  dampers  which  may  be  adjusted  by  a  positioning  damper 
motor  to  give  temperature  control. 

5  •  How  may  temperature  be  controlled  in  a  room  cooled  by  a  unit  cooler? 

Practice  indicates  that  a  thermostat  should  provide  for  the  automatic  operation  at  all 
hours  of  the  fan  and  control  valve  on  the  refrigeration  source,  but  that  there  be  a  manual 
switch  to  enable  the  fan  to  operate  continuously  during  occupancy. 

6  •  How  may  temperature  control  be  obtained  in  a  room  cooled  by  a  self- 
contained  mechanical  unit? 

The  fan  operation  may  be  controlled  by  a  manual  switch,  while  a  room  thermostat  in  con- 
junction with  a  solenoid  valve  may  regulate  the  flow  of  the  refrigerant  to  the  coil.  The 
thermostatic  circuit  might  be  operative  only  when  the  fans  are  running;  and  the  com- 
pressor might  be  controlled  by  refrigerant  pressure. 

256 


CHAPTER  14 — AUTOMATIC  CONTROL 


7  •  How  may  temperature  control  be  obtained  in  a  room  heated  by  an  auto- 
matically-fired warm  air  furnace? 

A  room  thermostat  might  control  the  combustion  unit;  and  a  limit  switch  in  the  top  of 
the  furnace  unit,  when  at  a  low  setting  of  its  control  might  operate  the  fan  whenever 
there  is  a  rise  of  temperature,  and  when  at  a  high  setting  of  its  control  it  might  shut  off 
the  combustion  unit.  A  room  humidity  control  operating  a  solenoid  valve  on  the  water 
supply  to  the  humidifier,  or  operating  a  relay  on  the  recirculating  pump  motor  to  the 
humidifier,  may  be  connected  in  parallel  with  the  fan  motor.  Humidification  may  be 
supplied  only  when  heat  is  supplied  and  when  the  humidity  control  acts  in  conjunction 
with  a  time  switch. 

8  •  How  may  humidity  be  controlled  in  a  unit  humidifier  for  a  steam  or  hot 
water  heating  plant? 

Since  heat  is  required  for  evaporation,  a  temperature  limit  switch,  preferably  of  the 
immersion  type,  may  be  placed  in  the  heating  supply  riser  to  cause  the  unit  to  be  in- 
operative when  heat  is  not  available.  A  room  humidity  control  will  operate  a  solenoid 
valve  on  the  water  supply  to  the  sprays.  Both  the  solenoid  valve  and  the  humidity 
control  may  be  electrically  wired  in  parallel  with  a  fan  motor,  and  be  subject  to  the 
temperature  limit  switch. 

9  •  Discuss  a  control  system,  including  control  of  humidity,  for  the  heating 
cycle  of  a  central  fan  system  of  air  conditioning. 

During  the  heating  cycle  it  is  necessary  to  vary  the  amount  of  outdoor  air  drawn  into  the 
system  in  accordance  with  the  temperature  of  that  air.  It  is  also  advisable  to  adjust  the 
volume  of  return  air  when  mixing  it  with  the  outdoor  air  so  that  the  resultant  mixture 
will  be  of  constant  volume  delivered  to  the  preheater  coils  at  some  predetermined  con- 
stant temperature. 

By  placing  a  temperature  controller  in  the  conditioner  just  ahead  of  the  preheating  coil, 
the  temperature  of  the  air  delivered  at  that  point  may  be  measured,  and  by  connecting 
this  controller  to  a  damper  motor  attached  to  the  intake  darriper  this  damper  can  be 
operated  by  a  temperature  variation  at  the  controller.  The  intake  damper  is  so  linked 
to  the  return  damper  that  the  combined  volume  of  air  delivered  through  the  ducts  of  the 
system  is  constant.  At  a  fall  in  outdoor  temperature,  this  arrangement  will  move  the 
intake  damper  to  a  closed  position  and  the  return  damper  to  an  open  position,  whereas 
the  reverse  will  hold  true  when  there  is  a  rise  in  outdoor  temperature. 
If  conditions  prevent  such  mechanical  linkage,  it  is  possible  to  use  two  damper  motors 
connected  so  they  are  operated  individually  but  in  inverse  ratio. 

The  operation  of  the  preheating  coils  should  be  dependent  upon  humidity  conditions  in 
the  occupied  spaces,  and  the  humidity  controller  should  be  installed  where  conditions 
are  representative  of  the  humidity  throughout  the  section,  because  air  leaving  the  pre- 
heating coils  is  immediately  passed  through  a  spray  where  it  becomes  saturated  with 
moisture.  If  the  air  is  cold,  it  will  absorb  so  little  moisture  that  when  it  is  delivered  to 
the  conditioned  spaces  its  relative  humidity  will  be  low.  When  the  compensated  hu- 
midity control  calls  for  additional  moisture,  the  steam  control  valve  in  the  preheater 
line  should  be  opened  to  allow  more  steam  to  flow  through  the  coils. 

Whenever  the  preheating  coils  are  being  heated  the  spray  should  be  in  operation,  but 
when  the  coils  are  cut  off  the  air  is  sufficiently  moist  and  the  spray  should  be  closed 
down.  This  necessitates  an  inter-connection  between  the  control  valve  on  the  pre- 
heater and  the  spray  pump  on  the  water  supply.  Water  is  supplied  to  the  spray  during 
the  heating  cycle  from  a  recirculating  water  tank  beneath  the  sprays. 

The  reheating  coil  determines  the  dry-bulb  temperature  of  the  delivered  air,  so  if  the 
conditioner  is  equipped  with  both  face  and  by-pass  dampers  on  this  coil  it  is  obvious  that 
these  dampers  should  be  controlled  by  a  thermostat  located  at  some  representative 
position  in  the  space  being  supplied  with  the  conditioned  air.  If  this  thermostat  is  in 
turn  connected  with  auxiliary  apparatus  which  will  vary  the  damper  settings,  it  will  be 
possible  to  pass  more  or  less  air  through  the  reheater  as  the  temperature  falls  or  rises. 

A  low-limit  temperature  control  might  also  be  mounted  in  the  discharge  duct  as  a 
precaution  against  blowing  cold  air  into  the  space.  Such  control  would  actuate  the 
dampers  of  the  reheater  when  the  duct  temperature  fell  below  a  predetermined  minimum 
regardless  of  the  demands  of  the  master  controller. 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

The  amount  of  steam  supplied  to  the  reheater  coils  should  be  a  function  of  the  position 
of  the  dampers.  If  the  face  dampers  are  closed  no  heat  is  required,  and  to  conserve 
steam  suitable  interconnection  between  the  damper  motor  and  the  control  valve  should 
be  made  in  order  that  this  valve  will  close  whenever  the  damper  valve  is  closed.  By 
adding  modulating  auxiliary  apparatus  to  the  steam  valve,  it  may  be  made  to  operate 
proportionately  to  the  setting  of  the  dampers. 

10  •  What  is  the  relation  between  comfort,  economy,  and  the  use  of  tempera- 
ture controls? 

As  a  general  rule,  a  moderate  expenditure  for  control  equipment  can  be  justified  on  the 
basis  of  economy,  but  the  cost  of  a  complete  system  of  individual  room  control  can 
ordinarily  be  only  partly  so  justified  and  the  remainder  must  be  charged  to  convenience 
and  comfort.  There  are,  however,  many  types  of  systems  where  the  question  would  not 
arise,  for  without  complete  control  equipment  these  systems  would  be  unusable. 


258 


Chapter  15 

AIR  POULUTION 

Sources  of  Air  Pollution,  Effects  of  Air  Pollution  on  Health,  Pul- 
monary   Effects,    Occlusion    of   Solar   Radiation,    Industrial   Air 
Pollution,  Abatement  of  Atmospheric  Pollution,  Smoke  Abate- 
ment, Dust  and  Cinder  Abatement 

THIS  chapter  considers  the  hygienic  aspects  of  atmospheric  pollution 
and  the  methods  by  which  this  pollution  may  be  lessened.  Infor- 
mation concerning  the  cleaning  of  air  brought  into  buildings  for  ventilat- 
ing purposes  will  be  found  in  Chapter  16,  and  a  discussion  of  the  exhaust- 
ing of  dusts  and  toxic  gases  from  factories  and  industrial  plants  is  con- 
sidered in  Chapter  21. 

The  impurities  which  contribute  to  atmospheric  pollution  include 
carbon  from  the  combustion  of  fuels,  particles  of  earth,  sand,  ash,  rubber 
tires,  leather,  animal  excretion,  stone,  wood,  rust,  paper,  threads  of 
cotton,  wool,  and  silk,  bits  of  animal  and  vegetable  matter,  and  pollen. 
Microscopic  examination  of  the  impurities  in  city  air  shows  that  a  large 
percentage  of  the  particles  are  carbon.  (See  Fig.  1,  Chapter  16,  for  size 
of  impurities  in  air.) 

Dust,  Fumes,  Smoke 

The  most  conspicuous  sources  of  atmospheric  pollution  may  be 
arbitrarily  classified  according  to  the  size  of  the  particles  as  dusts,  fumes, 
and  smoke.  Dusts  are  particles  of  solid  matter  varying  from  1.0  to  150 
microns  in  size.  Fumes  include  particles  resulting  from  chemical  pro- 
cessing, combustion,  explosion,  and  distillation,  ranging  from  0.1  to  1.0 
micron  in  size.  Smoke  is  composed  of  fine  soot  or  carbon  particles,  less 
than  0.1  micron  in  size,  which  result  from  incomplete  combustion  of 
carbonaceous  materials,  such  as  coal,  oil,  tar,  and  tobacco.  In  addition  to 
carbon  and  soot,  smoke  contains  unconsumed  hydrocarbon  gases,  sulphur 
dioxide,  sulphuric  acid,  carbon  monoxide,  and  other  industrial  gases 
capable  of  injuring  property,  vegetation,  and  health. 

The  lines  of  demarcation  in  these  three  classifications  are  neither  sharp 
nor  positive,  but  the  distinction  is  descriptive  of  the  nature  and  origin  of 
the  particles,  and  their  physical  action.  Dusts  settle  without  appreciable 
agglomeration,  fumes  tend  to  aggregate,  smoke  to  diffuse.  Particles 
larger  than  one  micron  will  eventually  settle  out  by  gravitation ;  particles 
smaller  will  remain  in  suspension  as  permanent  impurities  unless'  they 
agglomerate  to  sizes  larger  than  one  micron. 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

Fly-Ash,  Cinders 

The  term  fly-ash  is  usually  applied  to  the  extremely  small  particles  of 
ash,  and  the  term  cinder  to  the  larger  particles  of  coke  and  ash  which  are 
discharged  with  the  gases  of  combustion  from  burning  coal. 

AIR  POLLUTION  AND  HEALTH 

Many  kinds  of  dusts  and  gases  are  capable  of  producing  pathological 
changes  which  may  cause  ill  health.  The  harmful  effects  depend  largely 
upon  the  chemical  and  physical  nature  of  the  impurities,  and  the  con- 
centration, length  of  time,  and  conditions  under  which  they  are  breathed. 
Dust  particles  must  be  minute  in  size  to  be  inhaled  at  all,  although  fairly 
large  particles  may  gain  access  to  the  upper  air  passages. 

The  human  body  possesses  remarkable  filtering  media  for  protecting 
the  lungs.  Small  hairs  which  line  the  nasal  passages,  and  a  multitude  of 
microscopic  hairs,  called  cilia,  In  the  epithelial  lining  in  the  bronchial 
tubes  intercept  many  of  the  dust  particles  before  they  reach  the  lungs. 

The  constant  inhalation  of  dusts  in  city  air  irritates  the  mucous  mem- 
branes of  the  nose,  throat,  and  lungs,  and  eventually  may  produce  dis- 
comfort and  a  series  of  minor  respiratory  disorders.  The  pigmented  lung 
of  the  city  dweller  is  an  example  of  the  pathological  change  produced  over 
a  period  of  years.  This  condition  may  be  of  no  clinical  importance,  but 
an  exaggeration  of  it  in  the  coal  miner  results  in  anthracosis  or  dark  spots 
on  the  lung  due  to  the  presence  of  pigment  in  the  lymph  channels  which 
impairs  the  functioning  of  the  lung  cells  under  stress. 

Effects  of  Solids 

Bronchitis  is  the  chief  condition  associated  with  exposure  to  thick  dust, 
and  follows  upon  inhalation  of  practically  any  kind  of  insoluble  and  non- 
colloidal  dust.  Atmospheric  dust  in  itself  cannot  be  blamed  for  causing 
tuberculosis,  but  it  appears  to  have  a  marked  influence  in  aggravating  the 
disease  once  it  has  started.  There  is,  however,  quite  reliable  evidence 
that  carbon  pigment,  one  of  the  atmospheric  dusts,  tends  to  wall  off  local 
tuberculosis  rather  than  to  further  its  spread. 

The  sulphurous  fumes  and  tarry  matter  in  smoke  are  probably  more 
dangerous  than  the  carbon.  In  foggy  weather  the  accumulation  of  these 
substances  in  the  lower  strata  may  be  such  as  to  cause  irritation  of  the 
eyes,  nose,  and  respiratory  passages,  leading  to  asthmatic  breathing  and 
bronchitis  and,  in  extreme  cases,  to  death.  The  Meuse  Valley  fog 
disaster  will  probably  become  a  classic  example  in  the  history  of  gaseous 
air  pollution.  Released  in  a  rare  combination  of  atmospheric  calm  and 
dense  fog,  it  is  believed  that  sulphur  dioxide  and  other  toxic  gases  from 
the  industrial  region  of  the  valley  caused  63  sudden  deaths,  and  injuries 
to  several  hundred  persons.  Physical  examination  showed  difficult 
breathing,  rapid  pulse,  cyanosis,  cardiac  dilation,  and  a  redness  and 
inflammation  of  the  mucosa  of  the  nose,  mouth,  throat,  trachea,  and 
bronchi. 

Carbon  monoxide  from  automobiles  and  from  chimney  gases  con- 
stitutes another  important  source  of  aerial  pollution  in  busy  cities. 
During  heavy  traffic  hours  and  under  atmospheric  conditions  favorable  to 
concentration,  the  air  of  congested  streets  is  found  to  contain  enough  CO 

260 


CHAPTER  15 — AIR  POLLUTION 


to  menace  the  health  of  those  exposed  over  a  period  of  several  hours, 
particularly  if  their  activities  call  for  deep  and  rapid  breathing.  In  open 
air  under  ordinary  conditions  the  concentration  of  CO  in  city  air  is 
believed  to  be  insufficient  to  affect  the  average  city  dweller  or  pedestrian. 

Occlusion  of  Solar  Radiation 

The  loss  of  light,  particularly  the  occlusion  of  solar  ultra-violet  light 
due  to  smoke  and  soot,  is  beginning  to  be  recognized  as  a  health  problem 
in  many  industrial  cities.  Measurements  of  solar  radiation  in  Baltimore1 
by  actinic  methods  show  that  the  ultra-violet  light  in  the  country  was 
50  per  cent  greater  than  in  the  city.  In  New  York  City2  a  loss  as  great  as 
50  per  cent  in  visible  light  was  found  by  the  photo-electric  cell  method, 

The  effect  of  air  pollution  on  the  health  of  city  dwellers  is  difficult  to 
determine,  owing  to  the  slowness  of  its  manifestations.  The  aesthetic  and 
economic  objections  to  air  pollution  are  so  definite,  and  the  effect  of  air- 
borne pollen  can  be  shown  so  readily  as  the  cause  of  hay  fever  and  other 
allergic  diseases,  that  means  and  expenses  of  prevention  or  elimination  of 
this  pollution  have  seemed  justifiable  to  the  public. 

AIR  POLLUTION  IN  INDUSTRY 

In  many  industrial  processes,  sufficient  amounts  of  dusts,  fumes,  and 
vapors  are  liberated  to  be  injurious  to  the  health  of  workers.  Some  dusts 
are  poisonous  (lead,  mercury,  arsenic,  manganese,  and  cadmium)  and 
some  act  as  irritants  (silica,  steel,  iron,  and  granite).  Certain  dusts  may 
produce  catarrhal  conditions  and  increase  susceptibility  to  such  diseases 
as  bronchitis,  pneumonia,  and  tuberculosis.  Silicious  dust  is  especially 
harmful  because  it  has  a  direct  damaging  action  upon  the  tissue  of  the 
lungs,  but  organic  dusts,  both  animal  and  vegetable  (hair,  pollen,  textile, 
and  fiber),  do  not  seem  to  affect  the  lungs  at  all,  although  they  may  cause 
considerable  discomfort  in  the  upper  respiratory  passages  to  persons 
sensitive  to  them. 

Industrial  gases  and  fumes  act  specifically  upon  the  mucous  mem- 
branes, the  lungs,  blood,  skin,  and  eyes.  Some  extremely  poisonous  gases 
act  after  very  short  exposures.  Among  these  are  carbon  monoxide, 
hydrogen  sulphide,  ammonia,  chlorine,  bromine,  arsine,  and  cyanogen. 

The  industrial  processes  which  liberate  harmful  substances  are  too 
manifold  and  the  effects  too  diverse  to  be  considered  here,  where  dis- 
cussion is  limited  to  the  commonest  and  most  serious  with  which  the 
ventilating  engineer  may  be  confronted,  namely,  carbon  monoxide,  lead, 
and  silica.  For  a  more  thorough  treatise  on  the  subject  reference  should 
be  made  to  books  by  Hamilton3,  Ro'senau4,  and  Henderson  and  Haggard5. 

Carbon  Monoxide  Poisoning 

Carbon  monoxide  is  a  common  form  of  poisonous  industrial  gas,  met 
with  in  mines,  foundries,  coke-oven  sheds,  garages,  and  houses.  Its  action 

1Effects  of  Atmospheric  Pollution  upon  Incidence  of  Solar  Ultra-Violet  Light,  by  J.  H,  Shrader,  M.  H. 
Coblentz  and  F.  A.  Korff  (American  Journal  qfPttblic  Health,  p.  7,  Vol.  19,  1929). 

^Studies  in  Illumination,  by  J.  E.  Ivea  (U.  S.  Public  Health  Service  Bulletin  No  197,  1930). 
'Industrial  Poisons  in  the  United  States,  by  Alice  Hamilton.  • 

'Preventive  Medicine  and  Hygiene,  by  Milton  J.  Roseaau. 
Noxious  Gases,  by  Y.  Henderson  and  H.  Haggard. 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

is  due  to  the  fact  that  the  combining  power  of  carbon  monoxide  with  the 
haemoglobin  of  the  red  blood  corpuscles  is  about  300  times  greater  than 
that  of  oxygen.  Since  the  resulting  stable  combination  destroys  the 
power  of  the  haemoglobin  to  unite  with  oxygen  in  the  lungs  and  to  supply 
it  to  the  tissues,  the  effects  are  due  to  lack  of  oxygen,  and  the  symptoms 
are  those  of  anoxemia,  namely,  dizziness,  headaches,  sleepiness,  fatigue, 
and,  in  extreme  cases,  paralysis  and  death.  The  dangerous  saturation 
level  of  the  blood  with  carbon  monoxide  is  about  50  per  cent.  Even  as 
little  as  0.07  per  cent  in  the  air  will  render,  in  half  an  hour,  one  quarter  of 
the  red  corpuscles  incapable  of  uniting  with  oxygen.  One  to  two  parts 
per  10,000  parts  of  air  is  set  as  a  safe  limit  of  pollution  which  may  be 
breathed  for  a  long  time  without  producing  perceptible  symptoms. 

Silicosis 

Silicosis  is  a  chronic  disease  of  the  lungs  which  results  from  the  local 
physio-chemical  action  of  hydrated  silica  upon  the  pulmonary  tissue, 
causing  progressive  lymphatic  fibrosis,  and  rendering  the  tissue  suscep- 
tible to  tuberculosis.  The  disease  is  slow  in  evolution,  requiring  usually  a 
number  of  years  of  exposure.  It  occurs  principally  among  granite 
workers,  sand  blasters,  metal  miners,  metal  polishers,  potters,  and  mill- 
stone workers. 

Lead  Poisoning 

Lead  poisoning  is  the  most  insidious  and  most  common  of  all  industrial 
diseases.  It  occurs  principally  among  lead  workers  and  smelters,  lead 
miners,  potters,  painters,  typesetters,  stereotypers,  plumbers,  and 
workers  with  glass,  gold  and  silver.  Lead,  in  practically  all  forms,  ^is  a 
cumulative  poison  which  is  absorbed  by  way  of  the  blood  stream,  chiefly 
from  the  respiratory  tract,  but  also  from  the  digestive  tract  and  from  the 
skin.  The  effect  may  be  either  an  acute  or  chronic  poisoning.  The 
principal  symptoms  are  colic,  constipation,  anemia,  headache,  anorexia,  a 
bluish  line  along  the  edges  of  the  gums,  rheumatic  pains,  and,  in  extreme 
conditions,  paralysis,  blindness,  insanity,  and  death. 

It  has  been  found6  that  2  mg  per  day  is  the  smallest  dose,  by  inhalation, 
which  in  the  course  of  years  may  result  in  le£d  poisoning.  Regular 
inhalation  during  the  usual  working  hours  of  air  containing  less  than 
0.2  mg  of  lead  per  cubic  meter  does  not  seem  to  produce  serious  lead 
poisoning  in  individuals  of  representative  industrial  groups7. 

Prevention 

The  prevention  of  industrial  hazards  from  dusts  and  poisonous  gases  is 
largely  a  ventilation  problem  consisting  of  keeping  the  impurities  in  air 
down  to  a  safe  concentration.  As  yet  there  are  no  generally  accepted 
standards  on  which  to  base  the  design  of  the  ventilation  equipment. 
Approximate  data  on  the  toxicity  of  various  gases  and  fumes  met  with  in 
industrial  establishments  are  given  in  Table  1.  Column  5,  giving  the 
maximum  allowable  concentrations  for  prolonged  exposures,  was  com- 
piled from  experiments  in  which  most  exposures  lasted  not  more  than  a 

•Lead  Poisoning,  by  Thomas  Morrison  Legge  (Journal  Royal  Society  Arts,  1929,  Vol.  77,  p.  1023). 
*What  is  a  Dangerous  Quantity  of  Lead  I>ust  in  Air,  by  C.  M.  Sails  (Industrial  Hygiene  Bulletin,  New 
York  State  Department  of  Labor,  1925). 

262 


CHAPTER  15 — AIR  POLLUTION 


week,  and  it  is  reasonable  to  assume  that  over  more  prolonged  exposures 
such  concentrations  would  cause  pernicious  effects. 

Much  is  known  concerning  the  physiological  and  pathological  effects 
induced  by  various  types  and  concentrations  of  atmospheric  pollutants. 
In  the  absence  of  an  accepted  standard  for  safe  breathing,  and  because  of 
the  slow,  cumulative  effects  of  certain  kinds  of  air  contaminants,  the 
best  procedure  is  the  periodic  medical  examination  of  individuals,  and  the 

TABLE  1.    TOXICITY  OF  GASES  AND  FUMES  IN  PARTS  PER  10,000  PARTS  OF  AIR* 


VAPOR  OR  GAS 

RAPIDLY 
FATAL 

MAXIMUM 
CONCENTRATION 

FOB  FROM 

Yi  TO  1  HOUR 

MAXIMUM 
CONCENTRATION 
FOR  1  HOUR 

MAXIMUM 
ALLOWABLE 
FOR  PROLONGED 
EXPOSURE 

Carbon  monoxide 

40 

15-20 

10 

1 

Carbon  dioxide  

800-1000 

Hydrocyanic  acid  

30 

1& 

y> 

1< 

Ammonia  

50-100 

25 

o 

Hydrochloric  acid  gas  

10-20 

^ 

Mo 

Chlorine 

10 

<l 

Mnn 

Hydrofluoric  acid  gas  

2 

Mo 

Ms 

Sulphur  dioxide.-  

4-5 

i/  1 

/2~~-*- 

Kft 

Hydrogen  sulphide 

10-30 

5-7 

2-3 

1 

Carbon  bisulphide.-,.  
Phosphene.  
Arsine               

"20" 
2K 

11 
4-6 
Vz 

5 
1-2 
1A 

y* 

Phosgene. 

Over  1A 

1A 

rs 

Nitrous  fumes  

2J4-7J^ 

i-iH 

•K 

Benzene  

Toluene  and  xylene  
Aniline 

190 
190 

31-47 
31-47 

1-1  y> 

Mft 

Nitrobenzene  .          

Moo 

Xnn 

Petrol 

243 

100-220 

Carbon  tetrachloride  .  . 

480 

240 

40 

16 

Chloroform  

250 

140 

50 

2 

Tetrachlorethane  —  
Trichlorethylene  - 

73 
370 

1J* 

Methyl  chloride.  
Methyl  bromide  

1500-3000 
200-400 

200-400 
20-40 

70 
10 

5-10 
2 

Lead  vapor.  

5-6 

^Original  data  compiled  by  Y.  Henderson  and  H.  Haggard.  (See  Noxious  Gases,  1927.)  Data  revised 
by  T.  M.  Legge.  (See  Lessons  Learned  from  Industrial  Gases  and  Fumes,  Institute  of  Chemistry  of  Great 
Britain  and  Ireland,  London,  1930.) 

routine  measurement  and  study  of  the  concentration  and  the  physical  and 
chemical  characteristics  of  the  dusts  to  which  those  individuals  are 
exposed. 

ABATEMENT  OF  SMOKE  AND  AIR  POLLUTION 

Successful  abatement  of  atmospheric  pollution  requires  the  combined 
efforts  of  the  combustion  engineer,  the  public  health  officer,  and  the 
public  itself.  The  complete  electrification  of  industry  and  railroads,  and 
the  separation  of  industrial  and  residential  communities  would  aid 
materially  in  the  effective  solution  of  the  problem. 

In  the  large  cities  where  the  nuisance  from  smoke,  dust  and  cinders  is 

263 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

the  most  serious,  limited  areas  obtain  some  relief  by  the  use  of  district 
heating.  The  boilers  in  these  plants  are  of  large  size  designed  and  oper- 
ated to  burn  the  fuel  without  smoke,  and  some  of  them  are  equipped  with 
dust  catching  devices.  The  gases  of  combustion  are  usually  discharged  at 
a  much  higher  level 'than  is  possible  in  the  case  of  buildings  that  operate 
their  own  boiler  plants. 

In  general,  time,  temperature  and  turbulence  are  the  essential  require- 
ments for  smokeless  combustion.  Anything  that  can  be  done  to  increase 
any  one  of  these  factors  will  reduce  the  quantity  of  smoke  discharged. 
Especial  care  must  be  taken  in  hand-firing  bituminous  coals.  (See 
Chapter  27.) 

Checker  or  alternate  firing,  in  which  the  fuel  is  fired  alternately  on 
separate  parts  of  the  grate,  maintains  a  higher  furnace  temperature  and 
thereby  decreases  the  amount  of  smoke. 

Coking  and  firing,  in  which  the  fuel  is  first  fired  close  to  the  firing  door 
and  the  coke  pushed  back  into  the  furnace  just  before  firing  again,  pro- 
duces the  same  effect.  The  volatiles  as  they  are  distilled  thus  have  to 
pass  over  the  hot  fuel  bed  where  they  will  be  burned  if  they  are  mixed  with 
sufficient  air  and  are  not  cooled  too  quickly  by  the  heat-absorbing  surfaces 
of  the  boiler. 

Steam  or  compressed  air  jets,  admitted  over  the  fire,  create  turbulence 
in  the  furnace  and  bring  the  volatiles  of  the  fuel  more  quickly  into  contact 
with  the  air  required  for  combustion.  These  jets  are  especially  helpful 
for  the  first  few  minutes  after  each  firing.  Frequent  firings  of  small 
charges  shorten  the  smoking  period  and  reduce  the  density.  Thinner 
fuel  beds  on  the  grate  increase  the  effective  combustion  space  in  the 
furnace,  supply  more  air  for  combustion,  and  are  sometimes  effective  in 
reducing  the  smoke  emitted,  but  care  should  be  taken  that  holes  are  not 
formed  in  the  fire.  A  lower  volatile  coal  or  a  higher  gravity  oil  always 
produces  less  smoke  than  a  high  volatile  coal  or  low  gravity  oil  used  in 
the  same  furnace  and  fired  in  the  same  manner. 

The  installation  of  more  modern  or  better  designed  fuel  burning  equip- 
ment, or  a  change  in  the  construction  of  the  furnace,  will  often  reduce 
smoke.  The  installation  of  a  Dutch  oven  which  will  increase  the  furnace 
volume  and  raise  the  furnace  temperature  often  produces  satisfactory 
results. 

In  the  case  of  new  installations,  the  problem  of  smoke  abatement  can 
be  solved  by  the  selection  of  the  proper  fuel-burning  equipment  and 
furnace  design  for  the  particular  fuel  to  be  burned  and  by  the  proper 
operation  of  that  equipment.  Constant  vigilance  is  necessary  to  make 
certain  that  the  equipment  is  properly  operated.  In  old  installations  the 
solution  of  the  problem  presents  many  difficulties,  and  a  considerable 
investment  in  special  apparatus  is  necessary. 

Legislative  measures  at  the  present  time  are  largely  concerned  with  the 
smoke  discharged  from  the  chimneys  of  boiler  plants.  Practically  all  of 
the  ordinances  limit  the  number  of  minutes  in  any  one  hour  that  smoke  of 
a  specified  density,  as  measured  by  comparison  with  a  Ringelmann  Chart 
(Chapter  40),  may  be  discharged.  , 

These  ordinances  do  not  cover  the  smoke  discharged  at  low  levels  by 
automobiles,  and,  although  they  have  been  instrumental  in  reducing  the 

264 


CHAPTER  15 — AIR  POLLUTION 


smoke  emitted  by  boiler  plants,  they  have,  in  many  instances,  increased 
the  output  of  chimney  dust  and  cinders  due  to  the  use  of  more  excess  air 
and  to  greater  turbulence  in  the  furnaces. 

Legislative  measures  in  general  have  not  as  yet  covered  the  noxious 
gases,  such  as  sulphur  dioxide  and  sulphuric  acid  mist,  which  are  dis- 
charged with  the  gases  of  combustion.  Where  high  sulphur  coals  are 
burned,  these  sulphur  gases  present  a  serious  problem. 

DUST  AND  CINDERS 

The  impurities  in  the  air  other  than  smoke  come  from  so  many  sources 
that  they  are  difficult  to  control.  Only  those  which  are  produced  in 
large  quantities  at  a  comparatively  few  points,  such  as  the  dust,  cinders 
and  fly-ash  discharged  to  the  atmosphere  along  with  the  gases  of  com- 
bustion from  burning  solid  fuel,  can  be  readily  controlled. 

Dusts  and  cinders  in  flue  gas  may  be  caught  by  various  devices  on  the 
market,  such  as  fabric  filters,  dust  traps,  settling  chambers,  centrifugal 
separators,  electrical  precipitators,  and  gas  scrubbers,  described  in  later 
paragraphs. 

The  cinder  particles  are  usually  larger  in  size  than  the  dust  particles; 
they  are  gray  or  black  in  color,  and  are  abrasive.  Being  of  a  larger  size, 
the  range  within  which  they  may  annoy  is  limited. 

The  dust  particles  are  usually  extremely  fine;  they  are  light  gray  or 
yellow  in  color,  and  are  not  as  abrasive  as  cinder  particles.  Being  ex- 
tremely fine,  they  are  readily  distributed  over  a  large  area  by  air  currents. 

The  nuisance  created  by  the  solid  particles  in  the  air  is  dependent  on 
the  size  and  physical  characteristics  of  the  individual  particles.  The 
difficulty  of  catching  the  dust  and  cinder  particles  is  principally  a  function 
of  the  size  and  specific  gravity  of  the  particles. 

Lower  rates  of  combustion  per  square  foot  of  grate  area  will  reduce  the 
quantity  of  solid  matter  discharged  from  the  chimney  with  the  gases  of 
combustion.  The  burning  of  coke,  coking  coal,  and  sized  coal  from  which 
the  extremely  fine  coal  has  been  removed  will  not  as  a  general  rule  produce 
as  much  dust  and  cinders  as  will  result  from  the  burning  of  non-coking 
coals  and  slack  coal  when  they  are  burned  on  a  grate. 

Modern  boiler  installations  are  usually  designed  for  high  capacity  per 
square  foot  of  ground  area  because  such  designs  give  the  lowest  cost  of 
construction  per  unit  of  capacity.  Designs  of  this  type  discharge  a 
large  quantity  of  dust  and  cinders  with  the  gases  of  combustion,  and  if 
pollution  of  the  atmosphere  is  to  be  prevented,  some  type  of  catcher  must 
be  installed. 

Dust  and  Cinder  Catchers8 

The  various  types  of  dust  and  cinder  catchers  available  today  can  be 
divided  into  six  general  classes: 

1.  Settling  chambers. 

2.  Dust  and  cinder  traps. 

3.  Centrifugal  separators. 


•See  Smoke  and  Dust  Abatement,  by  M.  D.  Engle  (A.S.H.V.E.  TRANSACTION,  Vol.  37, 1931). 

265 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

4.  Electrostatic  precipitators. 

5.  Gas  scrubbers. 

6.  Fabric  filters. 

The  selection  of  the  proper  type  of  catcher  calls  for  a  careful  study  of 
the  material  to  be  caught  and  the  draft  and  space  available.  After 
installation,  constant  vigilance  is  necessary  to  keep  the  catchers  in  proper 
working  condition  if  satisfactory  operation  is  to  be  obtained. 

If  possible,  the  dust  or  cinder  catcher  should  be  installed  on  the  inlet 
side  of  the  induced  draft  fans  because  the  dust  and  cinders  in  the  gases 
seriously  erode  the  wheels  of  the  fans,  the  inlet  connectioxis  and  the 
scrolls.  Where  the  induced  draft  fans  operate  at  high  tip  speeds  and  no 
catchers  are  installed,  it  is  not  uncommon  for  the  fans^to  require  major 
repairs  within  one  year  and  complete  replacement  within  five  years. 

Settling  Chambers 

Probably  the  oldest  form  of  dust  catcher  is  the  settling  chamber, 
which  generally  consists  of  a  large-sized,  gas-tight  space  into  which  the 
dust-laden  gases  are  discharged  before  being  delivered  to  the  chimney. 
The  velocity  of  the  gas  should  be  reduced  to  a  point  where  the  larger  and 
heavier  particles  will  be  precipitated  by  gravity.  For  good  operation,  the 
velocity  of  the  gas  should  be  reduced  to  a  maximum  of  2  f  ps.  The  bottoms 
of  the  chambers  should  be  provided  with  dump  plates  through  which  the 
collected  dust  can  be  removed.  Because  these  chambers  are  not  effective 
in  removing  the  finer  dust  particles  they  have  been  practically  superseded 
by  smaller  and  less  costly  devices. 

Traps,  Catchers,  Precipitators 

Various  types  of  traps  have  been  devised.  In  general  they  all  depend 
upon  breaking  the  gas  up  into  thin  ^strata  and  subjecting  those  thin 
strata  to  several  abrupt  changes  in  direction.  The  dust  is  thrown  out 
of  the  gas  stream  into  specially  shaped  pockets,  or  impinged  against  a 
roughened  surface.  The  trapping  pockets  are  drained  into  a  hopper 
below  with  a  small  quantity  of  gas  and  the  dust  settles  out  by  gravity  due 
to  the  low  velocity  in  the  hopper.  In  the  roughened  surface  type,  various 
sections  of  the  trap  are  closed  off  at  intervals  by  means  of  dampers  and 
the  dust  is  shaken  off  the  roughened  surface  into  a  hopper  below. 

These  devices  work  very  well  in  catching  large  size  dust  and  cinders  and 
trap  much  of  the  fine  dust.  They  have  been  used  most  extensively  on 
stoker-fired  installations.  They  have  the  advantages  of  low  pressure 
drop,  relatively  small  space  requirements,  and  low  first  cost. 

Centrifugal  catchers  obtain  separation  by  projecting  the  particles 
tangentially  out  of  the  gas  stream.  The  effectiveness  of  this  type  of 
catcher  varies  directly  as  the  specific  weight  of  the  dust  and  as  the  square 
of  the  tangential  velocity,  and  inversely  as  the  radius  of  rotation. 

Electrostatic  precipitators  are  used  for  catching  fine  dust.  These 
precipitators  consist  of  dust-tight  chambers  in  which  are  suspended  rein- 
forced concrete  slabs  on  about  10-in.  centers.  Between  the  slabs  are 
suspended  bare  metal  rods.  High- voltage  unidirectional  current^  is 
applied  to  the  reinforcing  rods  in  the  concrete  slabs  acting  as  positive 
electrodes,  the  bare  rods  acting  as  negative  electrodes.  The  dust-laden 

266 


CHAPTER  15 — AIR  POLLUTION 


gas  flows  horizontally  through  the  precipitator  and  the  dust  particles 
migrate  toward  the  concrete  slabs  to  which  they  adhere  and  then  fall  or 
are  scraped  off  into  the  dust  hoppers  below. 

Gas  Scrubbers 

Wet  scrubbers  have  been  used  for  many  years  for  removing  dust  from 
gases.  A  number  of  different  types  of  scrubbers  are  now  being  built  for 
removing  dust  from  boiler  flue  gases.  One  type  depends  upon  saturating 
the  gas  and  washing  the  dust  out  of  suspension  by  a  spray  of  water.  For 
best  results  with  this  type,  the  water  should  be  atomized  into  as  fine  a 
spray  as  possible. 

Another  type  depends  upon  splitting  the  gas  into  thin  strata  and 
subjecting  these  strata  to  a  number  of  abrupt  changes  in  direction, 
throwing  the  dust  against  the  wet  surfaces.  The  main  problem  in  develop- 
ing a  satisfactory  wet  dust  catcher  is  to  find  suitable  materials  of  con- 
struction that  will  resist  the  corrosive  action  of  the  wash  water  for  a 
reasonable  length  of  time. 

Fabric  Filters 

Filters  of  many  kinds  have  been  used  with  variable  success.  The 
filter  bags  are  made  of  cotton,  wool  or  asbestos  fabric.  The  fabrics  used 
in  these  filters  do  not  withstand  the  temperatures  at  which  gases  are 
usually  discharged  from  the  boilers,  and  hence  the  gases  must  be  cooled  by 
some  means.  Surface  coolers  or  water  sprays  can  be  used  for  reducing  the 
gas  temperatures. 

One  of  the  serious  objections  to  all  of  these  dust  catchers  is  the  relatively 
high  cost  of  installation  and  maintenance,  and  the  space  required  for 
installation. 

Disposal  of  Dust  and  Cinders 

Even  after  the  dust  and  cinders  have  been  caught,  the  disposal  of  the 
material  caught  presents  a  serious  problem.  The  cinders  discharged  with 
the  gases  from  stoker-fired  boilers  are  usually  very  high  in  carbon  and 
contain  from  50  to  80  per  cent  as  much  heat  per  pound  as  the  coal  which 
is  being  burned.  It  is  possible,  and  usually  economical,  to  burn  these 
cinders.  They  cannot  be  satisfactorily  mixed  with  the  coal  in  the  stoker 
hopper  but  they  can  be  blown  into  the  furnace  over  the  stoker  fuel  bed 
and  burned  satisfactorily.  If  a  sufficient  quantity  of  cinders  is  caught,  a 
small  unit  pulverizer  can  be  installed  to  prepare  them  for  burning  over 
the  stoker  fuel  bed.  The  same  pulverizer  can  be  used  for  coal  at  times  of 
peak  load  and  will  materially  increase  the  capacity  of  the  fuel-burning 
equipment  for  the  boiler  to  which  it  is  connected. 

No  satisfactory  market  has  been  developed  for  the  dust  caught  from 
pulverized  coal  installations,  but  the  possibilities  are  being  investigated 
and  it  seems  likely  that  in  the  future  this  material  will  have  a  market 
value  that  will  go  a  long  way  toward  paying  the  fixed  charges  on  the  cost 
of  catching  it. 

The  distribution  of  dust  in  the  gas  entering  and  leaving  the  dust  and 
cinder  catchers  is  not  uniform  and  is  different  in  practically  every  in- 

267 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

stallation,  and  varies  widely  with  changes  in  furnace  conditions.  In 
order  to  obtain  a  representative  sample  it  is  necessary  to  traverse  the 
inlet  and  outlet  of  the  catcher  with  a  sampling  tube  which  faces  into  the 
gas  flow.  The  velocity  of  the  gas  into  the  sampling  tube  must  be  the 
same  as  the  velocity  of  the  gas  in  the  duct  at  the  instant  the  sample  is 
taken.  The  swirls  and  eddy  currents  in  the  ducts  make  it  difficult  to 
obtain  consistent  readings,  but  if  the  test  is  conducted  by  some  one  of 
experience,  an  indication  of  the  approximate  efficiency  can  be  obtained. 

Nature's  Dust  Catcher 

Nature  has  provided  means  for  catching  solid  particles  in  the  air  and 
depositing  them  upon  the  earth.  A  dust  particle  forms  the  nucleus  for 
each  rain  drop  and  the  rain  picks  up  dust  as  it  falls  from  the  clouds  to  the 
earth.  In  fact,  without  dust  in  the  air  to  form  the  nuclei  for  rain  drops  it 
would  never  rain,  and  the  earth  would  be  continually  enveloped  in  a  cloud 
of  vapor. 

PROBLEMS  IN  PRACTICE 

1  •  What  is  a  micron? 

A  micron  equals  0.001  millimeter  or  approximately  Jisoo  in. 

2  •  Distinguish  between  dusts,  fumes,  and  smokes. 

Solid  particles  ranging  in  size  from  1.0  micron  to  150  microns  are  called  dusts. 

Particles  resulting  from  sundry  chemical  reactions  and  ranging  from  0.1  to  1.0  micron  in 

size  are  called  fumes. 

Carbon  particles  less  than  0.1  micron  in  size  which  generally  arise  from  the  incomplete 

combustion  of  such  materials  as  coal,  oil,  or  tobacco  are  called  smokes. 

3  •  What  are  some  of  the  more  important  physical  properties  of  these  various 
groups  of  foreign  bodies  which  are  of  importance  in  ventilation? 

In  slowly  moving  air,  dusts  tend  to  settle  out  by  gravity  without  agglomerating  to  form 
larger  particles;  fumes  have  the  tendency  to  form  larger  particles  which  will  settle  when 
they  attain  the  size  of  approximately  1.0  micron ;  while  smokes  tend  to  diffuse  and  remain 
in  the  air  as  permanent  impurities. 

4  •  Why  is  atmospheric  pollution  an  important  engineering  problem? 

a.  Certain  impurities,  when  present  in  too  great  concentrations,  cause  ill  health  or  even 
death. 

b.  High  concentrations  of  solids  occlude  solar  radiations. 

c.  Some  materials  cause  permanent  injury  to  parts  of  buildings,  as  sulphur  fumes  corrode 
exposed  metal. 

d.  Extra  cleaning  expense  is  incurred  in  dusty  localities. 

e.  Internal  combustion  engines  are  damaged  by  abrasive  dusts. 

5  •  How  may  the  hazards  of  dust-producing  industrial   operations   best  be 
curtailed? 

By  providing  mechanical  exhaust  ventilation  sufficient  to  keep  dust  concentration  at  a 
safe  level  (see  Table  1)  and  then  removing  foreign  bodies  to  reduce  the  pollution  of  out- 
side air. 

6  •  How  may  the  pollution  of  the  atmosphere  be  lessened? 

By  compelling  industrial  plants  to  install  dust  catching  and  smoke  controlling  devices. 
In  many  cities  the  domestic  heating  plant  is  one  of  the  most  serious  offenders,  but  these 

268 


CHAPTER  15 — AIR  POLLUTION 


plants  are  too  small  to  justify  the  installation  of  dust  catchers.  Public  education  in 
improved  firing  methods  would  be  of  considerable  help  in  this  field. 

7  •  Compare  the  dry  and  wet  types  of  dust  catchers. 

The  dry  types  are  very  effective  in  removing  the  larger  dust  particles  but  the  smaller 
particles  generally  pass  through  other  kinds  than  the  electric  precipitator,  The  dry 
types  also  require  considerable  space  and  therefore  sometimes  introduce  resistance  to 
the  flow  of  air.  The  wet  types  are  effective  in  removing  some  of  the  smaller  dusts  and  the 
water-soluble  gases.  The  principal  disadvantage  of  the  washer  is  its  short  life  caused 
by  the  corrosive  action  of  the  wash  water. 

8  •  What  size  particles  are  detrimental  to  health? 

While  fairly  large  particles  may  enter  the  upper  air  passages,  those  found  in  the  lungs 
are  seldom  more  than  10  microns  in  size,  and  comparatively  few  of  them  are  more  than 
5  microns.  It  is  agreed  that  particles  between  J^  and  2  microns  may  be  harmful;  some 
authorities  place  the  upper  limit  at  about  5  microns,  and  some  incline  to  extend  the 
lower  limit  to  0.1  of  a  micron. 

9  •  Is  the  shape  of  the  particle  of  any  significance? 

Hard  particles  with  sharp  corners  or  edges  have  a  cutting  effect  on  the  delicate  mucous 
membranes  of  the  upper  respiratory  tract  which  may  lower  the  resistance  of  the  nose  and 
throat  to  acute  infections.  This  is  aggravated  by  the  irritating  effects  of  some  chemical 
compounds  which  may  be  taken  in  with  the  air  and  which  act  to  reduce  resistance. 

10  •  What  are  the  principal  meteorological  effects  of  smoke  and  dust? 

a.  The  reduction  in  the  amount  of  light  received.     Measurements  have  shown  that 
visible  light  may  be  as  much  as  50  per  cent  less  intense  in  a  smoky  section  of  a  city  than 
in  a  section  that  is  free  from  smoke.    Ultra-violet  light  is  reduced  as  much  or  more,  and 
in  some  cases  is  cut  out  entirely  for  a  time. 

b.  Smoke  and  dust  aid  in  the  formation  and  prolongation  of  fogs.    City  fogs  accumulate 
smoke  and  become  darker  in  color  and  very  objectionable.    The  sun  requires  a  longer 
time  to  disperse  them,  and  when  the  water  is  evaporated,  there  is  a  rain  of  smoke  and 
soot  particles  that  have  been  entrained. 

11  •  Why  has  not  smoke  abatement  been  more  effective? 

Because  communities  have  not  been  made  sufficiently  aware  of  the  possibilities  of 
burning  high  volatile  fuels  smokelessly  and  of  separating  cinder  and  ash  from  the  stack 
gases  to  a  degree  that  will  prevent  a  nuisance. 

12  •  Is  the  abatement  of  dust  and  cinders  important? 

Yes.  Only  a  small  percentage  of  the  solid  emission  from  stacks  is  smoke,  in  the  accepted 
popular  sense;  the  remainder  is  fly-ash  and  cinders.  While  black  smoke  is  disagreeable 
and  its  tarry  matter  and  carbon  particles  soil  anything  with  which  they  come  in  contact, 
the  cinders  and  some  of  the  ash  are  hard  and  destructive.  They  also,  together  with 
dusts  from  industrial  processes,  make  up  the  hard,  sharp,  irritating,  air-borne  solids 
that  are  breathed  by  individuals  not  working  in  a  dusty  mill  or  factory. 

13  •  Are  air-borne  impurities  causative  factors  in  hay  fever,  bronchial  asthma, 
and  allergic  disorders? 

Yes.  Recent  medical  investigations  indicate  that  90  per  cent  of  seasonal  hay  fever  and 
40  per  cent  of  bronchial  asthma  are  caused  by  air-borne  pollens,  tree  dusts,  and  other 
allergic  irritants. 

14  •  Name  some  essential  requirements  for  the  smokeless  combustion  of  fuels. 

Time,  temperature,  and  turbulence.  A  study  of  these  factors  is  usually  of  value  in 
overcoming  a  smoke  nuisance. 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

15  •  What  is  the  Ringelmann  Chart  Method  of  comparing  smoke  densities? 

See  Chapter  40.  The  Ringelmann  Chart  consists  of  four  cards  ruled  with  lines  having 
different  degrees  of  blackness.  These  cards,  together  with  a  white  card  and  a  black  one, 
are  hung  in  a  horizontal  row  50  ft  from  the  observer.  At  this  distance  the  lines  become 
invisible  and  the  cards  appear  to  be  different  shades  of  gray,  ranging  from  white  to  black. 
The  observer,  by  matching  the  cards  against  the  shades  of  smoke  coming  from  a  stack,  is 
able  to  estimate  the  blackness  of  the  smoke  as  compared  with  the  chart. 


270 


Chapter  16 

AIR  CLEANING  DEVICES 

Requirements  of  an  Air  Cleaner,  Types,  Air  Washers  and  Scrubbers, 
Viscous    Type   Filters,    Dry   Air   Filters,    Air   Filter   Installations 

THE  removal  of  impurities  from  air  brought  into  a  building  for 
ventilating  or  air  conditioning  purposes  is  the  function  of  any  air 
cleaning  or  filtering  device.  These  impurities  include  carbon  (soot)  from 
the  incomplete  combustion  of  fuels  burned  in  furnaces  and  automobile 
engines,  particles  of  earth,  sand,  ash,  automobile  tires,  leather,  animal 
excretion,  stone,  wood,  rust  and  paper,  threads  of  cotton,  wool  and  silk, 
bits  of  animal  and  vegetable  matter,  bacteria  and  pollen.  Microscopic 
examination  shows  that  the  character  of  the  impurities  varies  with  the 
locality,  but  as  a  rule  carbon  forms  the  greater  part  of  them  while  the 
total  is  somewhat  proportional  to  the  state  of  industrial  activity  and  the 
wind  intensity.  Additional  information  on  sources  of  air  pollution  will 
be  found  in  Chapter  15. 

Observations  have  shown  that  practically  all  atmospheric  impurities 
are  less  than  5  microns  in  size.  (One  micron  equals  0.001  millimeter  or 
approximately  0.00004  in.)  The  size  and  composition  of  each  individual 
particle  determines  its  buoyancy  and  consequently  the  length  of  time  it 
will  remain  in  suspension.  The  chart,  Fig.  1,  shows  graphically  the  sizes 
of  impurities  found  in  the  air,  and  other  related  data. 

To  estimate  the  probable  dust  load  for  air  filter  installations,  the 
following  approximate  averages  of  atmospheric  dust  concentration  may 
be  used  (7000  grains  equal  1  Ib) : 

Rural  and  suburban  districts 0.2  to  0.4  grains  per  1000  cu  ft 

Metropolitan  districts 0.4  to  0.8  grains  per  1000  cu  ft 

Industrial  districts 0.8  to  1.5  grains  per  1000  cu  ft 

REQUIREMENTS  OF  AN  AIR  CLEANER 

To  fulfill  the  essential  requirements  of  clean  air,  an  air  cleaner  should: 

1.  Be  efficient  in  the  removal  of  harmful  and  objectionable  impurities  in  the  air,  such 
as  dust,  dirt,  pollens,  bacteria. 

2.  Be  efficient  over  a  considerable  range  of  air  velocities. 

3.  Have  a  low  frictional  resistance  to  air  flow;  that  is,  the  pressure  drop  across  the 
filter,  measured  in  inches  of  water,  should  be  as  low  as  possible. 

4.  Have  a  large  dust-holding  capacity  without  excessive  increase  of  resistance,  or 
have  ability  to  operate  so  as  to  keep  the  resistance  constant  automatically. 

5.  Be  easy  to  clean  and  handle,  or  dean  itself  automatically. 

ft.  Leave  the  air  passing  through  the  cleaner  free  from  entrained  moisture  or  charging 
liquids  used  in  the  cleaner. 

271 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


The  A.S.H.V.E.  Standard  Code  for  Testing  and  Rating  Air  Cleaning 
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rated  by  (1)  capacity  in  cubic  feet  of  air  handled  per  minute,  (2)  resistance 


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FIG.  1.    SIZES  AND  CHARACTERISTICS  OF  AIR-BORNE  SOLIDS 

in  inches  of  water  at  rated  capacity,  (3)  dust  arrestance,  the  percentage 
relationship  expressing  dust  removal  efficiency  at  rated  capacity,  (4) 
reconditioning  power,  the  energy  necessary  to  operate  the  mechanism  of 


lAdopted  1934  by  A.S.H.V.E.    See  Chapter  41. 


272 


CHAPTER  16 — AIR  CLEANING  DEVICES 


an  automatic  air  cleaning  device,  and  (5)  dust  holding  capacity,  the 
amount  by  weight  of  standard  dust  which  a  non-automatic  air  cleaning 
device  will  retain  before  reconditioning  is  necessary. 

TYPES  OF  AIR  CLEANERS 

According  to  the  Code,  the  following  four  classifications  are  given  the 
devices : 

Class  A.  Automatic  Type:  In  general  all  air  cleaning  devices  which  use  power  to 
automatically  recondition  the  filter  medium  and  maintain  a  non-  vary  ing  resistance  to 
air  flow. 

Class  B.  Low  Resistance  Non- Automatic  Type:  Air  cleaning  devices  for  warm  air 
furnaces,  unit  ventilating  machines  and  similar  apparatus  and  installations  in  which  a 
maximum  of  not  more  than  0.18  in.  water  gage  is  available  to  move  air  through  the  air 
cleaning  device. 

Class  C.  Medium  Resistance  Non- A  utomatic  Type:  Air  cleaning  devices  for  systems 
in  which  a  maximum  of  not  more  than  0.5  in.  water  gage  is  available  to  move  air  through 
the  air  cleaning  device. 

Class  D.  High  Resistance  Non- A  utomatic  Type:  Air  cleaning  devices  for  the  air 
intake  of  compressors,  internal  combustion  engines,  and  the  like,  where  a  pressure  of 
1.0  in.  or  more  water  gage  is  available  to  move  air  through  the  air  cleaning  device. 

Air  cleaners  may  be  also  classified  as  follows: 

1.  According  to  principle  of  air  cleaning. 

a.  Air  washers. 

b.  Viscous  air  filters. 

(1)  Unit  type. 

(2)  Automatic  type. 

c.  Dry  air  filters. 

2.  According  to  application. 

a.  For  central  fan  systems  of  ventilation  and  air  conditioning.     Filters  of  the 
automatic  or  semi-automatic  type  are  usually  recommended  and  are  installed 
in  a  central  plenum  chamber. 

b.  For  unit  ventilators.    Filters  of  viscous  unit  or  dry  type,  installed  at  inlet  of 
individual  units. 

c.  For  window  installations.     Self-contained  units  consisting  of  fan  and  filter, 
usually  dry  type ,  adapted  to  be  placed  in  the  ordinary  window. 

d.  For  warm-air  furnaces.    Unit  type  viscous  or  dry  filters  placed  in  small  plenum 
chamber  of  warm-air  house  heating  systems. 

e.  For  compressors  and  Diesel  engines.    Unit  type  viscous  or  dry  filters,  installed  at 
air  intake  of  compressors  and  Diesel  engines. 

f.  For  compressed  air  lines.    Unit  type  viscous  or  dry  filters. 

With  the  growing  congestion  of  large  cities  and  an  industrial  growth 
throughout  the  entire  country,  the  percentages  of  foreign  material  in  the 
air,  such  as  soot  or  carbon,  which  are  unaffected  by  an  air  washer  type  of 
air  cleaner,  have  increased.  This  has  brought  about  the  development  of 
the  viscous  and  dry  type  air  filters  which  are  part  of  many  ventilating  and 
air  conditioning  systems. 

AIR  WASHERS  AND  SCRUBBERS 

Information  on  air  washers  will  be  found  in  Chapter  11. 
Scrubbers  have  not  been  used  very  extensively  in  the  past  for  cleaning 

273 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

air  for  ventilating  purposes.  However,  new  types  have  been  developed 
which  appear  to  have  possibilities  for  cases  where  the  air  to  be  cleaned  is 
extremely  dirty  or  where  a  higher  degree  of  cleanliness  is  desired  than  can 
be  obtained  with  an  air  washer. 

VISCOUS  TYPE  FILTERS 

The  principle  of  air  cleaning  used  in  viscous  filters  is  that  of  adhesive 
impingement.  Dust  and  dirt  in  the  air,  especially  soot  and  carbons,  are 
trapped  and  retained  by  successive  impingements  on  coated  surfaces. 
While  the  arrangements  of  filtering  media  and  the  kind  of  materials  used 
are  almost  unlimited,  there  are  certain  rather  definite  requirements  for  a 
practical  commercial  filter. 

Investigations  in  this  country  and  abroad  demonstrate  that  the  first 
impingement  of  dust  laden  air  on  a  viscous  coated  surface  removes  about 
60  per  cent  of  the  dust,  the  next  impingement  takes  60  per  cent  of  what 
then  remains — that  is,  24  per  cent — and  the  next  impingement  removes 
9.6  per  cent.  To  secure  maximum  efficiency,  it  is  necessary  to  divide  the 
air  into  innumerable  fine  streams,  as  the  more  intimately  and  freely  the 
air  is  brought  into  contact  with  the  viscous-coated  media  the  better  will 
be  the  cleaning. 

The  binding  liquid  used  with  viscous  filters  should  have  the  following 
properties : 

1.  Its  surface  tension  should  be  such  as  to  produce  a  homogeneous  film-like  coating 
on  the  filter  medium. 

2.  The  viscosity  should  vary  only  slightly  with  normal  changes  of  temperature. 

3.  It  should  be  germicidal  in  its  action  to  prevent  the  development  of  mold  spores 
and  bacteria  on  the  filter  media. 

4.  The  liquid  should  flow  freely  at  low  temperatures. 

5.  Evaporation  should  not  exceed  1  per  cent. 

6.  It  should  be  fireproof. 

7.  It  should  be  odorless. 

Viscous  Unit  Filters 

In  the  unit  type  viscous  filter,  the  filtering  media  are  arranged  in  units 
of  convenient  size  to  facilitate  installation,  maintenance,  and  cleaning. 
Each  unit  consists  of  an  interchangeable  cell  or  replaceable  filter  pad  and 
a  substantial  frame  which  may  be  bolted  to  the  frames  of  other  like  units 
to  form  a  partition  between  the  source  of  dusty  air  and  the  fan  inlet. 
The  necessary  washing,  draining,  and  recharging  equipment  should  be 
installed  near  each  group  of  unit  filters,  with  hot  water  and  sewer  con- 
nections provided. 

To  secure  greater  dust  holding  capacity  and  a  practically  constant 
resistance  and  air  volume,  the  filter  media  are  usually  placed  in  the 
direction  of  air  flow,  with  progressively  finer  filter  densities  determined 
by  the  percentage  of  dust  impinged.  This  arrangement  provides  relatively 
large  spaces  for  the  collection  of  dirt  in  the  front  of  the  filter  where  the 
bulk  of  the  dust  is  taken  out  without  undue  increase  in  resistance,  while 
at  the  back  of  the  filter  the  openings  are  smaller  to  secure  high  efficiency 
in  the  removal  of  the  finer  dust  particles. 

The  resistance  of  a  well-designed  unit  filter  of  the  adhesive  impinge- 

274 


CHAPTER  16 — AIR  CLEANING  DEVICES 


merit  type  usually  depends  upon  the  velocity  at  which  the  air  is  handled 
and  upon  whether  the  unit  is  clean  or  dirty.  The  cleaning  efficiency  ^of 
the  unit  is  usually  highest  after  it  has  accumulated  a  certain  portion  of  its 
maximum  load  of  dirt  because  some  dust  collected  in  the  cell  acts  as  an 
efficient  medium  for  the  further  seizing  of  solids  from  the  air.  By  periodi- 
cally cleaning  a  predetermined  number  of  cells,  the  resistance  and  capacity 
of  a  built-up  filter  may  be  held  at  any  desired  figure.  The  frequency  of 
cleaning  any  unit  filter  installation  depends  upon  the  dust  concentration 


0.30  & 


4 


12 


14 


16 


6  8  10 

•Hfy  of  Dusf  7  oz. 

FIG.  2.    CHART  SHOWING  CHANGE  IN  RESISTANCE  DUE  TO  DUST  ACCUMULATION 

0.40 


700         750         800        850        900 
Cubic  Feei  of  Air-ThroiKjh  Fitter  per  Minule 


950        1000 


FIG.  3.    RESISTANCE  TO  AIR-FLOW  OF  A  TYPICAL  UNIT  Am  FILTER 

of  air  being  cleaned,  and  on  the  amount  of  dirt  which  can  be  accumulated 
in  the  filter  medium  without  causing  excessive  resistance. 

Filters  consisting  of  inexpensive  frames  of  cardboard  or  similar  material 
filled  with  viscous-coated  glass  wool  or  steel  wool  are  available.  Because 
of  their  construction  these  units  may  be  discarded  when  dirty  and  replaced 
with  new  units  at  relatively  little  expense.  They  are  used  in  general 
ventilation  work  and  with  warm  air  furnaces  and  other  installations  where 
first  cost  and  low  resistance  to  air  flow  are  essential.  The  operating 
characteristics  of  these  units  conform  in  general  with  those  of  the  rigid 
frame  type. 

Viscous  Automatic  Filters 

The  principle  of  air  cleaning  used  in  the  viscous  automatic  filters  is 
the  same  as  in  the  unit  filters.  The  removal  of  the  accumulated  dust, 

275 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

however,  is  done  automatically  instead  of  by  hand.  The  automatic  clean- 
ing and  recoating  of  these  filters  is  based  on  the  principle  that  the  viscous 
fluid  itself  will  perform  the  cleaning  function,  thereby  eliminating  a  sepa- 
rate washing  agent.  The  dust  collected  by  the  filter  thus  is  deposited 
finally  in  the  bottom  of  the  viscous  fluid  reservoir  from  which  it  may  be 
removed  by  different  methods,  depending  on  the  design  of  the  filter. 

There  are  three  general  types  of  automatic  filters.  They  are  differentiated 
from  each  other  according  to  the  process  of  self-cleaning  and  renewing 
of  the  viscous  coating  used  by  each  type,  as  follows: 

1.  The  filter  medium  has  the  form  of  an  endless  curtain  suspended  vertically,  with  its 
lower  portion  submerged  in  a  viscous  fluid  reservoir.  The  curtain  rotates  slowly  through 
this  bath,  thus  performing  the  cleaning  and  recoating  of  the  filter  medium. 

2.  The  filter  screen  is  arranged  in  the  form  of  shelves  or  cylinders,  and  the  viscous 
fluid  is  flushed  through  all  parts  of  the  medium  in  a  direction  opposite  to  the  air  flow, 

3.  The  filter  medium  is  arranged  vertically  and  is  stationary.  The  viscous  fluid  is 
flushed  from  above  over  the  medium,  while  the  air  flow  is  stopped. 


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FIG.  4.    MAINTENANCE  CHART  FOR  UNIT  TYPE  Viscous  FILTERS 

The  washing  and  renewing  process  in  automatic  filters  usually  is  inter- 
mittent. It  is  accomplished  by  an  electric  motor  or  by  other  motive 
power  and  is  controlled  by  manual  or  by  automatic  timing  devices.  The 
operating  cycle  is  of  a  predetermined  frequency  and  should  be  so  timed 
as  to  insure  a  constant  static  pressure  drop  across  the  filter.  The  customary 
resistance  to  air  flow  is  i^-in.  water  gage  at  an  air  velocity  of  500  fpm, 
measured  at  the  filter  entrance.  Automatic  viscous  filters  are  made  up  in 
units  which  are  delivered  either  fully  assembled  or  in  parts  to  be  assem- 
bled at  the  point  of  installation, 

DRY  AIR  FILTERS 

Dry  air  filters,  in  which  dust  is  impinged  upon  or  filtered  through 
screens  made  of  felt,  cloth,  or  cellulose,  are  available  in  various  types. 
These  filters  require  no  adhesive  liquid,  but  depend  on  the  straining  or 
screening  action  of  the  filtering  medium.  Because  of  the  close  texture 

276 


CHAPTER  16 — AIR  CLEANING  DEVICES 


of  the  filtering  media  used  in  most  of  the  dry  filters,  the  surface  velocity, 
or  velocity  of  the  air  entering  the  media,  ranges  between  10  and  50  fpm, 
depending  on  the  nature  and  texture  of  the  fabric.  This  necessitates  a 
relatively  large  screen  surface,  and  the  filter  media  are  usually  arranged 
in  the  form  of  pockets  to  bring  the  frontal  area  within  customary  space 
requirements. 

As  in  viscous  unit  filters,  an  average  constant  resistance  and  air  volume 
may  be  obtained  by  periodic  reconditioning  or  renewal  of  the  filter 
screens.  Since  some  materials  suitable  for  dry  filtering  media  are  affected 
considerably  by  moisture  which  tends  to  cause  a  rapid  increase  in  resis- 
tance, they  should  be  treated  or  processed  to  minimize  the  effect  of 
changes  in  humidity. 

Filters  using  felt  and  similar  materials  as  filter  media  depend  upon 
vacuum  cleaning  for  reconditioning.  A  special  nozzle,  operated  from  a 
portable  or  stationary  vacuum  cleaner,  is  shaped  to  reach  all  parts  of  the 
filter  pockets.  Permanent  filter  media  should  be  capable  of  withstanding 
repeated  vacuum  cleanings  without  loss  in  dust  removal  efficiency. 
While  most  dry  filters  are  cleaned  by  replacing  an  inexpensive  filter  sheet, 
the  useful  life  of  these  sheets  often  may  be  lengthened  by  vibrating  or 
vacuum  cleaning. 

INSTALLATION  METHODS 

The  published  performance  data  for  all  air  filters  are  based  on  straight 
through  unrestricted  air  flow.  Filters  should  be  installed  so  that  the  face 
area  is  at  right  angles  to  the  air  flow  whenever  possible.  Eddy  currents 
and  dead  air  spaces  should  be  avoided  and  air  should  be  distributed 
uniformly  over  the  entire  filter  surface,  using  baffles  or  diffusers  if  neces- 
sary. 

The  most  important  requirements  of  a  satisfactory  and  efficiently 
operating  air  filter  installation  are: 

1.  The  filter  must  be  of  ample  size  for  the  amount  of  air  it  is  expected  to  handle.    Aii 
overload  of  10  to  15  per  cent  is  regarded  as  the  maximum  allowable.    When  air  volume  is 
subject  to  increase,  a  larger  filter  should  be  installed. 

2.  The  filter  must  be  suited  to  the  operating  conditions,  such  as  degree  of  air  clean- 
liness required,  amount  of  dust  in  the  entering  air,  type  of  duty,  allowable  pressure  drop, 
operating  temperatures,  and  maintenance  facilities. 

3.  The  filter  type  should  be  the  most  economical  for  the  specific  application.    The 
first  cost  of  the  installation  should  be  balanced  against  depreciation  as  well  as  expense 
and  convenience  of  maintenance. 

The  following  recommendations  apply  to  filters  and  washers  installed 
with  central  fan  systems: 

1.  Duct  connections  to  and  from  the  filter  should  change  size  or  shape  gradually  to 
insure  even  air  distribution  over  the  entire  filter  area. 

2.  Sufficient  space  should  be  provided  in  front  as  well  as  behind  the  filter  to  make  it 
accessible  for  inspection  and  service.  A  distance  of  two  feet  may  be  regarded  as  the 
minimum. 

3.  Access  doors  of  convenient  size  should  be  provided  in  the  sheet  metal  connections 
leading  to  and  from  the  filters. 

4.  All  doors  on  the  clean  air  side  should  be  lined  with  felt  to  prevent  infiltration  of 
unclean  air.  All  connections  and  seams  of  the  sheet  metal  ducts  oh  the  clean  air  side 
should  be  as  air-tight  as  possible. 

277 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

5.  Electric  lights  should  be  installed  in  the  chamber  in  front  of  and  behind  the  air  filter. 

6.  Air  washers  should,  whenever  possible,  be  installed  between  the  tempering  and 
heating  coils  to  protect  them  from  extreme  cold  in  winter  time. 

7.  Filters  installed  close  to  air  inlet  should  be  protected  from  the  weather  by  suit- 
able louvers,  in  front  of  which  a  large  mesh  wire  screen  should  be  provided. 

8.  Filters  should  have  permanent  indicators  to  give  a  warning  when  the  filter  re- 
sistance reaches  too  high  a  value. 

REFERENCES 

Testing  and  Rating  of  Air  Cleaning  Devices  Used  for  General  Ventilation  Work,  by 
Samuel  R.  Lewis  (A.S.H.V.E.  Journal  Section,  Heating,  Piping  and  Air  Conditioning, 
May,  1933). 

Fundamental  Principles  in  the  Design  of  Dry  Air  Filters,  by  Otto  Wechsberg 
(A.S.H.V.E.  Journal  Section,  Heating,  Piping  and  Air  Conditioning,  April,  1933). 

Operation  and  Maintenance  of  Air  Filters,  by  W.  G.  Frank  (Heating,  Piping  and  Air 
Conditioning,  May,  1931). 

Size  and  Characteristics  of  Air-Borne  Impurities,  by  W.  G.  Frank  (Heating,  Piping 
and  Air  Conditioning,  January,  1932). 

Determining  the  Quantity  of  Dust  in  Air  by  Impingement,  by  F.  B.  Rowley  and 
John  Beal  (A.S.H.V.E.  TRANSACTIONS,  Vol.  35,  1929). 

A  Study  of  Dust  Determinators,  by  F.  B.  Rowley  and  John  Beal  (A.S.H.V.E.  TRANS- 
ACTIONS, Vol.  34,  1928). 

Design  and  Application  of  Oil-Coated  Air  Filters,  by  H.  C.  Murphy  (A.S.H.V.E. 
TRANSACTIONS,  Vol.  33,  1927). 

Determining  the  Efficiency  of  Air  Cleaners,  by  A.  M.  Goodloe  (A.S.H.V.E.  TRANS- 
ACTIONS, Vol.  30,  1924). 


PROBLEMS  IN  PRACTICE 

1  •  What  is  meant  by  air  filter  performance  characteristics? 

The  factors  that  determine  the  performance  of  an  air  filter,  which  are: 

(1)  efficiency  in  dust  removal,  (2)  operating  resistance,  (3)  dust  holding  capacity.  In  a 
properly  designed  filter  these  factors  are  balanced  to  obtain  the  desired  characteristics 
for  a  given  application.  Since  the  requirements  vary  for  different  kinds  of  air  cleaning 
service,  it  is  necessary  to  have  filters  of  different  types  to  meet  the  various  conditions. 

2  •  What  are  the  advantages  of  viscous  filters? 

The  principal  advantage  of  the  viscous  filter  is  its  large  dust  holding  capacity.  The  dust 
accumulation  is  distributed  through  the  depth  of  the  filtering  medium  rather  than  upon 
the  surface  as  in  the  dry  types,  which  makes  it  possible  for  viscous  filters  to  handle 
heavy  dust  concentrations  without  excessive  resistance.  Since  its  efficiency  and  resis- 
tance are  based  on  maximum  air  velocities  of  from  300  to  500  ft  per  minute  through  the 
filter,  the  viscous  filter  consumes  the  minimum  amount  of  space  for  a  given  air  volume. 

3  •  What  are  the  advantages  of  dry  filters? 

Dry  filters  are  more  efficient  in  the  removal  of  fine  dust  particles  from  the  air,  and  some 
types  will  eliminate  even  as  much  as  60  per  cent  of  the  smoke  particles.  Dry  filters  also 
are  easily  and  conveniently  maintained  by  vacuum  cleaning,  vibrating,  or  renewing  the 
filtering  medium.  _  i 

4  •  If  an  air  washer  is  used  for  cooling  and  humidity  control  in  an  air  con- 
ditioning system,  is  a  filter  needed? 

An  air  filter  is  desirable  in  conjunction  with  an  air  washer  because  of  the  large  amount  of 
soot  in  the  air  which,  due  to  its  greasy  and  amorphous  nature,  is  not  readily  trapped  in 

278 


CHAPTER  16 — AIR  CLEANING  DEVICES 


an  air  washer.  Filters  should  be  placed  between  the  washer  and  the  air  intake  so  that 
all  the  dirt  will  be  collected  at  one  point  to  simplify  maintenance,  to  protect  all  the 
equipment  in  the  system,  and  to  prevent  contamination  of  the  water  used  in  the  washer. 

5  •  Is  an  air  filter  needed  with  an  extended  surface  type  heat  exchanger? 

An  air  filter  is  essential  with  an  extended  surface  heat  exchanger  in  order  to  maintain  its 
efficiency,  for  without  this  protection  dust  particles  will  adhere  to  the  exposed  surfaces, 
and  gradually  build  up  a  deposit  to  the  point  where  the  efficiency  will  be  impaired  and  the 
resistance  increased  by  restricting  the  air  passage. 

6  •  What  is  the  proper  location  of  a  filter  in  relation  to  the  fan? 

A  filter  will  operate  equally  well  whether  placed  on  the  suction  or  discharge  side  of  the 
fan.  It  has  become  standard  practice,  however,  to  locate  the  filter  on  the  fan  inlet  side 
because  there  it  has:  (1)  simpler  duct  connections,  (2)  reduced  static  pressure  losses, 
(3)  more  even  air  distribution  over  the  entire  filter  area.  Where  an  exceptionally  high 
efficiency  in  dust  removal  must  be  maintained,  it  is  often  advisable  to  place  the  filter  on 
the  discharge  side  of  the  fan  so  there  can  be  no  infiltration  of  unclean  air. 

7  •  What  instruments  and  apparatus  are  required  for  determining  the  pollen 
concentration  in  air  by  means  of  the  settling  method? 

A  microscope  with  a  field  of  know  area  and  a  glass  slide  coated  with  a  viscous  material. 

8  •  Describe  the  procedure  for  determining  the  pollen  concentration  in  air  by 
means  of  the  settling  method. 

A  glass  slide  coated  with  a  viscous  material  is  placed  for  a  period  of  24  hours  in  a  hori- 
zontal position  in  the  atmosphere  to  be  tested.  The  slide  is  then  removed  and  placed 
under  the  microscope,  and  pollen  counts  are  made  of  approximately  25  fields  over  the 
area  of  the  glass  slide.  Having  determined  the  count  over  a  definite  area,  as  for  example, 
1  sq  cm,  and  finding  the  settling  rate  of  the  average  particles  from  the  chart,  Fig.  1,  the 
concentration  in  parts  per  cubic  yard  can  be  calculated. 

9  •  The  resistance  to  ah*  flow  of  a  unit  air  filter  is  found  to  be  0.4  in.  of  water. 
The  volume  of  air  passing  through  the  filter  is  1000  cfm  at  a  velocity  of  200  fpm. 
What  would  be  the  filter  area  required  in  order  to  reduce  the  pressure  drop 
across  the  filter  from  0.4  in.  of  water  to  0.16  in.  of  water? 

Referring  to  Fig.  3:  The  resistance  is  substantially  proportional  to  the  square  of  the 
velocity,  or 

Q  =  It 
R«        V,} 

0.4  2002 


0.16  722 

F22  =  16,000 
F2  -  126.5  fpm 
Q  =  AV 
1000  =  126.5  A 


The  filter  area  would  be  increased  from  5  sq  f  t  to  7.91  sq  ft. 

10  •  A  ventilating  system  complete  with  filters  has  a  fan  which,  when  operating 
at  400  rpm  and  delivering  air  at  1  in.  of  water  total  static  pressure,  requires  an 
input  of  3  horsepower.  After  the  system  operates  for  a  time,  the  pressure  drop 
across  the  filter  caused  by  the  clogging  action  of  the  collected  dust  and  dirt 
increases  from  0.1  in.  of  water  to  0.4  in.  of  water.  To  maintain  the  original 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

rate  of  air  delivery  with  the  increased  static  pressure,  at  what  speed  must  the 
fan  be  run  and  what  horsepower  will  be  required? 

Static  pressure  after  clogging  of  filter  =  1  -j-  (0.4  —  O.lj  =  1.3  in.  of  water. 

The  static  pressure  varies  as  the  square  of  the  fan  speed.    Therefore,  if  X  is  the  fan  speed 
after  the  static  pressure  increases: 

1.3 


1          V  400 

X  =  456  rpm. 

The  horsepower  varies  as  the  cube  of  the  fan  speed.    Therefore,  if  Y  is  the  horsepower 
after  the  static  pressure  increases: 

456  \3 


__ 
3         V  400  / 

F  =  4.44  horsepower. 

To  maintain  the  original  rate  of  air  delivery  with  the  increased  static  pressure,  the  fan 
speed  must  be  increased  from  400  to  456  rpm,  and  the  horsepower  from  3  to  4.44. 


280 


Chapter  17 

FANS  AND  MOTIVE  POWER 

Performance,  Fan  Efficiency,  Characteristic  Curves,  Selection  of 
Fans,  Controls,  Designation  of  Fans,  Motive  Poiver,  Electric  Power 

FANS  are  used  for  producing  air  flow  except  where  positive  displace- 
ment is  required,  in  which  case  compressors  or  rotary  blowers  are 
used.  Fans  are  classified  according  to  the  direction  of  air  flow  as  (1) 
axial  flow  or  propeller  type  if  the  flow  is  parallel  with  the  axis,  and  (2) 
radial  flow  or  centrifugal  type  if  the  flow  is  parallel  with  the  radius  of 
rotation. 

Axial  flow  fans  are  made  with  various  numbers  of  blades  of  a  variety 
of  forms.  The  blades  may  be  of  uniform  thickness  (sheet  metal),  either 
flat  or  cambered,  or  may  be  of  varying  thickness  of  so-called  aerofoil 
section  (airplane  propeller  type).  Where  an  axial  flow  fan  is  intended  for 
operation  at  comparatively  high  pressures  the  hub  sometimes  is  enlarged 
in  the  form  of  a  disc  and  the  fan  is  known  as  a  disc  fan. 

Radial  flow  or  centrifugal  fans  include  steel  plate  fans,  pressure  blowers, 
cone  fans,  and  the  so-called  multiblade  fans.  All  the  foregoing  types  have 
variations  which  may  be  obtained  by  modification  of  the  proportions  or 
change  in  the  curvature  and  angularity  of  the  blades.  The  angularity  of 
the  blades  determines  the  operating  characteristics  of  a  fan:  a  forward 
curved  blade  is  found  in  a  fan  having  slow  speed  operating  characteristics, 
while  a  backward  curved  blade  is  found  in  a  fan  having  high  speed 
operating  characteristics. 

A  wide  variation  exists  in  the  demands  which  have  to  be  met  by  fan 
installations.  A  fan  may  be  required  to  move  large  quantities  of  air 
against  little  or  no  resistance  or  it  may  be  required  to  move  small  quanti- 
ties against  high  resistances.  Between  these  two  extremes  innumerable 
specific  requirements  must  be  met.  In  general,  fans  of  all  types  in  each 
general  class  can  be  made  to  perform  the  same  duty,  although  mechanical 
difficulties,  noise  or  lack  of  efficiency  may  limit  the  use  to  one  or  another 
type.  The  most  common  field  of  service  for  fans  of  the  propeller  type  is  in 
moving  air  against  moderate  resistances,  especially  where  no  long  ducts 
or  heavy  friction  must  be  overcome  and  where  noise  is  not  objectionable, 
whereas  centrifugal  fans  are  commonly  employed  for  operation  at  the 
comparatively  higher  pressures  and  where  extreme  quietness  is  necessary, 

PERFORMANCE  OF  FANS 

Fans  of  all  types  follow  certain  laws  of  performance  which  are  useful  in 
determining  the  effect  of  changes  in  the  conditions  of  operation.  These 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

laws  apply  to  installations  comprising  any  type  of  fan,  any  given  piping 
system  and  constant  air  density,  and  are  as  follows: 

1.  The  air  capacity  varies  directly  as  the  fan  speed. 

2.  The  pressure  (static,  velocity,  and  total)  varies  as  the  square  of  the  fan  speed. 

3.  The  power  demand  varies  as  the  cube  of  the  fan  speed. 

Example  1.  A  certain  fan  delivers  12,000  cfm  at  a  static  pressure  of  1  in.  of  water 
when  operating  at  a  speed  of  400  rpm  and  requires  an  input  of  4  hp.  If  in  the  same 
installation  15,000  cfm  are  desired,  what  will  be  the  speed,  static  pressure,  and  power? 


Speed  =  400  X  j  »  500  rpm 

/  ^oox  2 
Static  pressure  =  1  X  f  TTJA  )    —  1-56  in. 

Power  =  4  X  (g?)*  =  7.81  hp 

When  the  density  of  the  air  varies  the  following  laws  apply  : 

4.  At  constant  speed  and  capacity  the  pressure  and  power  vary  directly  as  the 
density. 

Example  2.  A  certain  fan  delivers  12,000  cfm  at  70  F  and  normal  barometric  pressure 
(density  0.07495  Ib  per  cubic  foot)  at  a  static  pressure  of  1  in.  of  water  when  operating  at 
400  rpm,  and  requires  4  hp.  If  the  air  temperature  is  increased  to  200  F  (density  0.06018 
Ib)  and  the  speed  of  the  fan  remains  the  same,  what  will  be  the  static  pressure  and 
power? 

Static  pressure  =  1  X  0*07495  ~  0-80  in- 


5.  At  constant  pressure  the  speed,  capacity  and  power  vary  inversely  as  the  square 
root  of  the  density. 

Example  3.  If  the  speed  of  the  fan  of  Example  2  is  increased  so  as  to  produce  a  static 
pressure  of  1  in.  of  water  at  the  200  F  temperature,  what  will  be  the  speed,  capacity, 
and  power? 


Capacity  -  12,000  X       -_  =  13,392  cfm  (measured  at  200  F) 
0.06018 


6.  For  a  constant  weight  of  air: 

(a)  The  speed,  capacity,  and  pressure  vary  inversely  as  the  density. 

(b)  The  horsepower  varies  inversely  as  the  square  of  the  density. 

Example  4-  If  the  speed  of  the  fan  of  the  previous  examples  is  increased  so  as  to 
deliver  the  same  weight  of  air  at  200  F  as  at  70  F,  what  will  be  the  speed,  capacity, 
static  pressure,  and  power? 


Capacity  =  12,000  X  =  14M5  cfm  (measured  at  200  F) 

282 


CHAPTER  17 — FANS  AND  MOTIVE  POWER 


Static  pressure  =  1   X  -n'f^rr-t~^  ~  1-25  in. 
U.UoULS 


FAN  EFFICIENCY 

The  efficiency  of  a  fan  may  be  defined  as  the  ratio  of  the  power  required 
in  moving  the  air  to  the  power  input  to  the  fan.  The  work  done  in 
moving  the  air  may  be  computed  on  the  basis  of  either  the  static  or  the 
total  pressure.  When  the  static  pressure  is  used  in  the  computation  it  is 
assumed  that  this  represents  the  useful  pressure  and  that  the  velocity 
pressure  is  lost  in  the  piping  system  and  in  the  air  which  leaves  the  system. 
Since  in  most  installations  a  higher  velocity  exists  at  the  fan  outlet  than 
at  the  point  of  delivery"  into  the  atmosphere,  some  of  the  velocity  pressure 
at  the  fan  outlet  may  be  utilized  by  conversion  to  static  pressure  within 
the  system,  but  owing  to  the  uncertainty  of  friction  losses  which  occur  at 
the  places  where  changes  in  velocity  take  place,  the  amount  of  velocity 
pressure  which  is  actually  utilized  is  seldom  known,  and  the  static  pressure 
alone  may  best  represent  the  useful  pressure. 

The  efficiency  based  upon  static  pressure  is  known  as  the  static  efficiency 
and  may  be  expressed  as  follows: 

St  t*      ffi  *         i  =  cfm  X  static  pressure  in  inches  of  water  . 

lency        6369  X  power  input  expressed  in  units  of  746  watts         (  ' 

Different  fans  may  develop  the  same  capacity  against  the  same  static 
pressure  and  with  the  same  power  input,  and  therefore  operate  at  the 
same  static  efficiency,  while  maintaining  different  outlet  velocities.  Where 
a  high  outlet  velocity  is  desirable  or  can  be  utilized  effectively,  the  static 
efficiency  fails  to  be  a  satisfactory  measurement  of  the  performance.  In 
many  applications  of  propeller  fans,  air  is  circulated  without  encountering 
resistance  and  no  static  pressure  is  developed.  The  static  efficiency  is 
zero  and  its  calculation  is  meaningless.  Because  of  such  situations  where 
the  static  efficiency  fails  to  indicate  the  true  performance,  many  engineers 
prefer  to  base  the  calculation  of  efficiency  upon  the  total  or  dynamic 
pressure.  This  efficiency  is  variously  known  as  the  total,  dynamic,  or 
mechanical  efficiency,  and  may  be  expressed  as  follows: 

T  t  I    ffi  *  —  cfm  X  total  pressure  in  inches  of  water  >«. 

iotal  efficiency  -  6359  x  ^^^  input  expressed  In  units  of  746  watts  (  ' 

CHARACTERISTIC  CURVES 

In  the  operation  of  a  fan  at  a  fixed  speed  the  static  and  total  efficiencies 
vary  with  any  change  in  the  resistance  which  is  imposed.  With  different 
designs  the  peak  of  efficiency  occurs  when  the  fans  deliver  different  per- 
centages of  their  wide-open  capacity.  Variations  in  efficiency  accompany 
variations  in  pressures  and  power  consumption  which  are  characteristic  of 
the  individual  designs  and  which  are  influenced  particularly  by  the  shape 


1See  Standard  Test  Code  for  Disc  and  Propeller  Fans,  Centrifugal  Fans  and   Blowers,  Edition  of 
1932. 

283 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

and  angularity  of  the  blades.    Such  variations  in  pressure,  power,  and 
efficiency  are  shown  by  characteristic  curves. 

Characteristic  curves  of  fans  are  determined  by  tests  performed  in 
accordance  with  the  Standard  Test  Code  for  Disc  and  Propeller  Fans, 
Centrifugal  Fans  and  Blowers2  as  adopted  by  the  AMERICAN  SOCIETY  OF 
HEATING  AND  VENTILATING  ENGINEERS  and  the  National  Association  of 
Fan  Manufacturers.  The  results  of  tests  are  plotted  in  different  ways :  the 
abscissae  may  be  the  ratio  of  delivery,  assuming  full  open  discharge  as 
100  per  cent,  and  the  ordinates  may  be  static  pressure,  dynamic  pressure, 
horsepower  and  efficiency.  Pressures  may  be  expressed  in  per  cent  of  the 
maximum  pressure  in  the  manner  shown  in  the  illustrations  in  this 


40  50  60 

Per  Cent  of  Wide  Open  Volume 

FIG.  1.    OPERATING  CHARACTERISTICS  OF  AN  AXIAL  FLOW  FAN 


chapter,  but  in  engineering  calculations  they  are  sometimes  expressed  in 
proportion  to  the  pressures  due  to  the  peripheral  velocity. 

It  should  be  noted  that  characteristic  curves  of  fan  performance  are 
plotted  for  a  constant  speed.  Some  variation  in  values  of  efficiency  may 
occur  at  different  speeds  but  such  variation  is  usually  slight  within  a  wide 
range  of  speeds.  Fans  of  similar  design  but  of  different  size  will  also  show 
some  difference  in  efficiency.  The  proportions  of  the  housing  also  affect 
the  performance.  As  a  rule  a  narrow  fan  of  large  diameter  shows  a  higher 
efficiency  than  one  of  greater  width  and  smaller  diameter.  For  a  number 
of  designs  using  blades  of  certain  shapes  the  proportion  of  the  width  to  the 
diameter  is  so  definitely  established  by  the  service  for  which  the  fan  is 
intended  that  little  variation  in  efficiency  occurs,  but  in  other  designs, 
particularly  that  which  uses  straight  radial  blades,  the  efficiency  may 
vary  over  a  wide  range  depending  on  whether  the  dimensions  are  suitable 
for  a  fan  intended  for  ordinary  ventilating  purposes  or  for  a  pressure 
blower.  Figs.  1  to  4  show  characteristic  curves  for  different  types  of  fans 


*A.S.H.V.E.  TRANSACTIONS,  Vol.  29,  1923.    Amended  June,  1931. 

284 


CHAPTER  17 — FANS  AND  MOTIVE  POWER 


using  blades  of  various  shapes,  but  without  reference  to  the  design  of 
housing  employed.  The  efficiency  curves  are  therefore  not  serviceable 
for  making  rigid  comparisons  of  efficiencies  obtainable  with  blades  of  the 
various  shapes  but  are  intended  merely  to  show  reasonable  values  and 
more  particularly  to  show  the  manner  in  which  variations  occur  with 
changes  in  fan  capacity. 

Axial  flow  fan  characteristics  are  indicated  by  Figs.  1  and  2.  These 
fans,  when  properly  designed,  have  a  satisfactory7  efficiency  at  low 
resistance,  comparing  favorably  in  this  respect  with  centrifugal  fans. 
They  are  low  in  cost  and  economical  in  operation  and  occupy  relatively 
little  space.  Although  this  type  of  fan  can  operate  against  considerable 


30  40  50  60  70 

Per  Cent  of  Wide  Open  Volume 


90 


FIG.  2.    OPERATING  CHARACTERISTICS  OF  AN  AIRPLANE  PROPELLER  FAN 

resistance,  the  noise^  often  becomes  objectionable,  so  that  it  does  not 
always  compare  favorably  with  centrifugal  fans  for  such  service.  With 
most  of  the  designs  which  employ  blades  of  uniform  thickness  the  power 
increases  rapidly  with  an  increase  in  resistance. 

The  curves  (Fig.  1)  show  the  rapid  reduction  in  capacity  and  increase  in 
power  as  the  resistance  increases.  The  low  efficiency  when  overcoming 
heavy  resistance  is  due  to  the  low  speed  of  the  blades  near  the  hub  as 
compared  to  the  relatively  high  peripheral  or  tip  speed.  The  air  driven  by 
the  blade  area  near  the  rim  can  pass  back  through  the  less  effective  blade 
area  at  the  hub  more  easily  than  it  can  overcome  the  duct  resistance. 

Fig.  2  shows  the  performance  of  the  airplane  propeller  fan  in  which  the 
blades  are  similar  in  shape  to  those  of  an  airplane  propeller  but  of  varying 
number  according  to  the  pressure  to  be  developed.  This  fan  usually 
operates  at  a  higher  speed  than  does  the  former  type  of  propeller  fan,  and 
with  a  different  power  characteristic,  the  power  remaining  fairly  constant 
throughout  the  range  of  pressures,  being  somewhat  less  at  the  higher  than 
at  the  lower  pressures.  The  flatness  of  the  pressure  curve  indicates  the 
advantage  of  this  type  of  fan  in  preventing  overloading  of  motors  where 
fluctuations  in  pressure  occur.  Variations  in  the  diameter,  width,  pitch, 

285 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

camber,  and  the  thickness  of  the  blades  provide  a  considerable  degree  of 
flexibility  in  design,  so  that  the  peak  of  total  efficiency  may  be  made  to 
occur  at  wide-open  volume  or  at  various  percentages  of  that  volume. 

Another  advantage  of  this  type  of  axial  flow  fan  is  its  low  resistance  to 
air  passage  when  standing  still.  There  are  some  installations  in  which 
such  a  characteristic  is  desirable. 

The  straight  blade  (paddle-wheel)  or  partially  backward  curved  blade 
type  of  fan  is  practically  obsolete  for  ventilation.  Its  use  is  largely  con- 
fined to  such  applications  as  conveyors  for  material,  or  for  gases  con- 
taining foreign  material,  fumes  and  vapors.  The  open  construction  and 
the  few  large  flat  blades  of  these  wheels  render  them  resistant  to  corrosion 
and  tend  to  prevent  material  from  collecting  on  the  blades.  This  type  of 
fan  has  a  good  efficiency,  but  the  power  steadily  increases  as  the  static 


Slio 


40  50  60  70 

Per  Cent  of  Wide  Open  Volume 


80 


90 


100 


FIG.  3.    OPERATING  CHARACTERISTICS  OF  A  FAN  WITH  BLADES  CURVED  FORWARD 

pressure  falls  off ,  which  requires  that  the  motor  be  selected  with  a  moder- 
ate reserve  in  power  to  take  care  of  possible  error  in  calculation  of  duct 
resistance. 

The  forward  curved  multiblade  fan  is  the  type  most  commonly  used  in 
heating  and  ventilating  work,  as  it  has  a  low  peripheral  speed,  a  large 
capacity,  and  is  quiet  in  operation.  The  point  of  maximum  efficiency  for 
this  fan  occurs  near  the  point  of  maximum  static  pressure.  The  static 
pressure  drops  consistently  from  the  point  of  maximum  efficiency  to  full 
open  operation.  Fig.  3  shows  that  this  type  of  fan  will  have  both  a  high 
and  a  low  delivery  for  a  given  static  pressure  at  constant  speed.  The 
power  curve  rises  continually  from  low  to  peak  capacity,  but  if  reasonable 
care  is  exercised  in  figuring  resistance  there  is  no  danger  of  overloading 
the  motor. 

The  outstanding  characteristics  of  the  full  backward  curve  multiblade 
type  fan  are  the  steep  pressure  curves,  the  non-overloading  power  curve, 
and  the  high  speed.  (See  Fig.  4.)  This  fan  operates  at  a  peripheral  speed 
of  approximately  250  per  cent  of  the  forward  curve  multiblade  type  for 

286 


CHAPTER  17 — FANS  AND  MOTIVE  POWER 


like  results.  The  pressure  curves  begin  to  drop  at  very  low  capacity  and 
continue  to  fall  rapidly  to  full  outlet  opening.  The  steep  pressure  curves 
tend  to  produce  constant  capacity  under  changing  pressures.  Where 
wide  fluctuations  in  demand  occur,  this  type  of  fan  is  desirable  to  prevent 
overloading  of  motors.  The  maximum  power  requirement  occurs  at 
about  the  maximum  efficiency.  Consequently  a  motor  selected  to  carry 
the  load  at  this  point  will  be  of  sufficient  capacity  to  drive  the  fan  over  its 
full  range  of  capacities  at  a  given  speed.  The  high  speed  of  this  type 
makes  it  adaptable  for  direct  connected  electric  motor  drives.  The  high 
speed  may  necessitate  somewhat  heavier  construction  and  more  operating 
attention  or  service.  The  dimensional  bulk  for  a  given  duty  often  is 
150  per  cent  of  that  of  a  forward  curve  multiblade  type  fan. 

Between  the  extremes  of  the  forward  and  the  full  backward  curve  blade 
type  centrifugal  fans  a  number  of  modified  designs  exist,  differing  in  the 


20  30  40  50  60 

Per  Cent  of  Wde  Open  Volume 


FIG.  4.    OPERATING  CHARACTERISTICS  OF  A  FAN  WITH  BLADES  CURVED  BACKWARD 

angularity  or  in  the  shape  of  the  blades.  Common  among  these  designs 
are  the  straight  radial  blade  type,  the  radial  tip  type,  and  the  double 
curve  blade  type  with  a  forward  angle  at  the  heel  and  a  slight  backward 
angle  at  the  tip  of  the  blade.  Characteristic  curves  of  these  types  show 
varying  degrees  of  resemblance  to  the  curves  of  Figs.  3  and  4,  according 
to  the  degree  of  similarity  to  one  or  the  other  of  the  two  designs  of  fan 
considered. 

SELECTION  OF  FANS 

The  following  information  is  required  to  select  the  proper  type  of  fan ; 

1.  Cubic  feet  of  air  per  minute  to  be  moved. 

2.  Static  pressure  required  to  move  the  air  through  the  system. 

3.  Type  of  motive  power  available. 

4.  Whether  fans  are  to  operate  singly  or  in  parallel  on  any  one  duct. 
&  What  degree  of  noise  is  permissible. 

6»  Nature  of  the  load,  such  as  variable  air  quantities  or  pressures. 

287 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

Knowing  the  requirements  of  the  system,  the  main  points  to  be  con- 
sidered for  fan  selection  are  (1)  efficiency,  (2)  speed,  (3)  noise,  (4)  size  and 
weight,  and  (5)  cost. 

In  order  to  facilitate  the  choice  of  apparatus,  the  various  fan  manu- 
facturers supply  fan  tables  or  curves  which  usually  show  the  following 
factors  for  each  size  of  fan  operating  against  a  wide  range  of  static 
pressures: 

1.  Volume  of  air  in  cubic  feet  per  minute  (68  F,  50  per  cent  relative  humidity, 
0.07488  Ib  per  cubic  foot). 

2.  Outlet  velocity. 

3.  Revolutions  per  minute. 

4.  Brake  power. 

5.  Tip  or  peripheral  speed. 

6.  Static  pressure. 

The  most  efficient  operating  point  of  the  fan  is  usually  shown  by  either 
bold-face  or  italicized  figures  in  the  capacity  tables. 

Fans  for  Ventilation  and  for  Cooling  Systems 

Two  important  factors  in  selecting  fans  for  ventilating  systems  are 
efficiency  (which  affects  the  cost  of  operation)  and  noise.  First  cost  and 
space  available  are  secondary.  The  fans  should  be  selected  to  operate 
at  maximum  efficiency  without  noise.  Because  noise  in  a  ventilating 
system  is  irritating  and  a  cause  for  complaint,  fans  must  be  selected  of 
proper  size  in  order  to  reduce  it  to  a  minimum.  Noise  may  be  caused  by 
other  factors  than  the  fan,  namely,  high  velocity  in  the  duct  work, 
unsatisfactory  location  of  the  fan  room,  improper  construction  of  floors 
and  walls,  and  poor  installation.  Where  noise  is  chargeable  directly  to 
the  fan,  it  is  caused  either  by  excessive  peripheral  speeds,  or  the  fan  is  of 
insufficient  size.  It  should  be  remembered,  however,  that  the  tip  speed 
required  for  a  specified  capacity  and  pressure  varies  with  the  type  of 
blade,  and  that  a  tip  speed  which  may  be  excessive  for  the  forward 
curved  type  is  not  necessarily  so  for  the  backward  or  slightly  backward 
type.  A  noisy  fan  usually  is  one  which  is  operated  at  a  point  considerably 
beyond  maximum  efficiency. 

For  a  given  static  pressure  there  is  a  corresponding  outlet  velocity  and 
peripheral  speed  wherein  maximum  efficiency  is  obtained.  If  a  fan  is 
selected  to  operate  at  this  point,  the  cost  of  operation  and  the  noise  can 
be  held  within  control. 

To  aid  in  selecting  fans  as  near  as  possible  to  the  point  of  maximum 
efficiency,  there  are  listed  in  Tables  1  and  2  for  each  static  pressure  cor- 
responding outlet  velocities  and  tip  speeds  which  will  give  satisfactory 
results.  The  proper  tip  speed  for  a  given  static  pressure  varies  with  the 
design  of  wheel  and  with  the  number  of  blades  or  vanes  in  the  wheel. 

Lower  outlet  velocities  than  those  listed  in  Table  1  may  be  employed, 
but  care  must  be  exercised  when  fans  of  the  forward  curved  type  are  used 
to  avoid  selecting  a  fan  for  operation  below  its  useful  range.  The  useful 
range  of  the  fans  of  Table  2  extends  over  the  full  length  of  the  per- 
formance curve. 

In  exhaust  ventilating  systems  where  the  air  column  moves  toward  the 

288 


CHAPTER  17 — FANS  AND  MOTIVE  POWER 


fan,  noise  due  to  the  higher  tip  speeds  and  outlet  velocities  will  not  be 
so  readily  transmitted  back  through  the  air  column  to  the  building  as 
when  the  air  column  is  moving  toward  the  rooms.  Therefore  higher 
outlet  velocities  may  be  used,  but  this  will  be  at  the  expense  of  increased 
horsepower. 

Amply  large  fans  should  always  be  used  for  both  exhaust  and  supply 
systems,  as  there  may  be  and  usually  is  leakage  despite  the  most  careful 
workmanship,  necessitating  the  delivery  of  more  air  at  the  fans  than  is 
exhausted  from  or  supplied  through  the  openings  in  the  various  rooms. 

Long  runs  of  distributing  ducts,  heaters,  and  air  washers  require 
definite  increments  of  the  total  pressure  which  a  supply  fan  in  a  venti- 
lating system  must  overcome.  These  static  pressures  should  be  con- 
sidered when  selecting  the  fan  characteristics,  speed,  and  power. 


TABLE  1. 


GOOD  OPERATING  VELOCITIES  AND  TIP  SPEEDS  FOR  FORWARD  CURVED 

MULTIBLADE   VENTILATING  FANS 


STATIC  PRESSURE 


OUTLET  VELOCITY 


TIP  SPEED 


INCHES  OP  WATER 

FEET  PEE  MINUTE 

FEET  PER  MINUTE 

M 

1000-1100 

1520-1700 

%                            looo-iioo 

1760-1900 

H                                        1000-1200 

1970-2150 

& 

1100-1300 

2225-2450 

1200-1400 

2480-2700 

% 

1300-1600 

2660-2910 

1 

1500-1800 

2820-3120 

JLM 

1600-1900 

3162-3450 

1J^ 

1800-2100 

3480-3810 

1H 

1900-2200 

3760-4205 

2 

2000-2400 

4000-4500 

2M 

2200-2600 

4250-4740 

2H 

2300-2600 

4475-4970 

3 

2500-2800 

4900-5365 

Fans  picked  within  the  limits  of  Table  1  will  operate  close  to  the  point 
of  maximum  efficiency.  No  attempt  has  been  made  to  select  these  limits 
for  quiet  operation,  since  this  is  a  relative  term  and  varies  with  the  type 
and  location  of  the  installation. 

The  connection  of  a  fan  to  a  metallic  duct  system  should  be  made  by 
canvas  or  a  similar  flexible  material  so  as  to  prevent  the  transmission  of 
fan  vibration  or  noises.  Where  noise  prevention  is  a  factor  the  fan  and  its 
driver  should  have  floating  foundations. 

Fans  for  Drying 

Both  axial  flow  and  centrifugal  types  of  fans  are  used  for  drying  work. 
Propeller  fans  are  well  adapted  to  the  removal  of  moisture-laden  air  when 
operating  against  low  resistance  and  when  handling  air  at  low  tempera- 
tures. Motors  on  these  fans  usually  are  of  the  fully-enclosed  moisture- 
proof  types  so  that  saturated  air  or  air  containing  foreign  material  will 
not  injure  the  motors. 

Unit  heaters  employing  axial  flow  fans  are  widely  used  in  the  drying 
field.  In  drying,  these  fans  may  be  used  with  unit  heaters  where  not 
too  much  duct  work  is  required  and  where  air  is  to  be  delivered  against 

289 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


pressure,  since  the  noise  developed  from  the  high  peripheral  speed  of  these 
fans  is  not  ordinarily  objectionable  in  process  work. 

Centrifugal  fans  of  the  multiblade  type  generally  are  selected  to  supply 
air  for  drying,  as  they  are  capable  of  delivering  large  volumes  of  air 
against  all  pressures  likely  to  be  encountered. 

Belt  driver!  fans  usually  are  to  be  preferred  to  direct-connected  fans 
since  efficient  motor  speeds  do  not  usually  coincide  with  efficient  fan 
speeds.  Replacement  of  a  standard  motor  is  quick  and  easy  if  it  is  belted. 

Wherever  drying  is  done  throughout  the  year  and  where  air  require- 
ments change  as  the  drying  conditions  change,  the  drying  can  be  speeded 
up  or  reduced  through  control  of  the  fan  capacity.  This  may  be  done  by 
changing  the  fan  speed  or  by  varying  the  outlet  area  with  dampers.  A 
throttled  outlet  reduces  the  volume  and  reduces  the  power. 

Due  to  the  low  speeds  of  forward  curved  multiblade  or  paddle-wheel 
type  fans,  these  can  be  direct-connected  to  reciprocating  steam  engines, 

TABLE  2.  GOOD  OPERATING  VELOCITIES  AND  TIP  SPEEDS  FOR  MULTIBLADE  VENTILATING 
FANS  WITH  BACKWARD  TIPPED  AND  DOUBLE  CURVED  BLADES 


STATIC  PRESSURE 
INCHES  of  WATER 

OUTLET  VELOCITY 
FEET  PER  MINUTE 

TIP  SPEED 
FEET  PER  MINUTE 

H 

800-1100 

2600-3100 

B/S 

800-1150 

3000-3500 

jl 

900-1300 

3400-4000 

H 

1000-1500 

3800-4500 

% 

1100-1650 

4200-5000 

% 

1200-1750 

4500-5300 

1                       :                1200-1900 

4800-5750 

1M 

1300-2100 

5300-6350 

1H 

1400-2300 

5750-6950 

iH 

1500-2500 

6200-7550 

2 

1600-2700 

6650-8050 

2J4 

1700-2800 

7050-8550 

2H 

1800-2950 

7450-9000 

3 

2000-3200 

,  8200-9850 

ctnd  the  exhaust  steam  from  the  engines  may  be  used  in  the  heating 
apparatus.  In  selecting  engine  driven  fans  for  drying  processes,  where  a 
large  quantity  of  exhaust  steam  is  used  in  the  heaters r  a  smaller  fan  and 
greater  power  consumption  may  be  used,  because  power  economy  is  not 
essential  under  this  condition. 

Where  static  pressure  in  a  dryer  varies,  and  where  several  fans  must 
operate  in  parallel,  fans  are  to  be  preferred  which  have  a  continuously 
rising  pressure  characteristic,  such  as  is  given  by  backward-curved  or 
double-curved  blades.  This  type  of  fan  is  well  adapted  for  direct-con- 
nected motors  of  the  higher  speeds. 

Fans  far  Dust  Collecting  and  Conveying 

The  application  of  fans  for  handling  refuse,  dust,  and  fumes  generated 
by  machine  equipment  is  covered  in  Chapter  21.  Information  is  given 
regarding  the  methods  for  determining  air  quantities,  the  velocity  required 
for  carrying  various  materials  and  the  method  of  determining  maintained 

290 


CHAPTER  17 — FANS  AND  MOTIVE  POWER 


resistance  or  total  static  pressure  at  which  the  fan  is  to  operate.  The 
selection  of  a  proper  size  fan  is  at  times  governed  by  the  future  require- 
ments of  the  plant.  In  many  instances,  additional  future  capacity  is 
anticipated  and  should  be  provided  for. 

Having  determined  the  necessary  volume  of  air  and  the  maintained 
resistance  or  static  pressure  required,  the  proper  size  fan  may  be  selected 
from  the  fan  manufacturers'  performance  charts  or  capacity  tables.  The 
fan  chosen  should  be  the  size  that  will  provide  the  required  ultimate 
quantities  with  the  minimum  power  consumption. 

FAN  CONTROL 

Some  method  of  volume  control  of  fans  usually  is  desirable.  This  may 
be  done  by  varying  the  peripheral  velocity  or  by  interposing  resistance,  as 
by  throttling-dampers.  Both  methods,  since  they  reduce  the  volume  of 
air,  reduce  the  power  required.  In  many  installations  adjustments  of 
volume  are  desirable  during  varying  hours  of  the  day.  In  others  an 
increased  supply  of  air  in  summer  over  that  needed  for  winter  is  demanded. 
Experience  is  required  in  deciding  whether  speed-control  or  damper- 
control  shall  be  used  for  specific  cases.  Where  noise  is  a  factor,  it  may  be 
exceedingly  desirable  to  reduce  the  speed  at  times,  while  on  the  other 
hand,  any  fan  which  has  its  normal  speed  reduced  as  much  as  50  per  cent 
without  change  in  resistance  will  move  only  50  per  cent  of  the  air. 

DESIGNATION  OF  FANS 

Facing  the  driving  side  of  the  fan,  blower,  or  blast  wheel,  if  the  proper  direction  of 
rotation  is  clockwise,  the  fan,  blower,  or  blast  wheel  will  be  designated  as  clockwise. 
If  the  proper  direction  of  rotation  is  counter-clockwise,  the  designation  will  be  counter- 
clockwise.  (The  driving  side  of  a  single  inlet  fan  i&  considered  to  be  the  side  opposite 
the  inlet  regardless  of  tie  actual  location  of  the  drive.)8 

This  method  of  designation  will  apply  to  all  centrifugal  fans,  single  or  double  width, 
and  single  or  double  inlet.  Do  not  use  the  word  "hand,"  but  specify  ''clockwise"  or 
* '  counter-clockwise." 

The  discharge  of  a  fan  will  be  determined  by  the  direction  of  the  line  of  air  discharge 
and  its  relation  to  the  fan  shaft,  as  follows: 

Bottom  Tiorizontal:  If  the  line  of  air  discharge  is  horizontal  and  below  the  shaft. 
Top  horizontal:  If  the  line  of  air  discharge  is  horizontal  and  above  the  shaft. 
Up  blast:  If  the  line  of  air  discharge  is  vertically  up. 
Down  blast:  If  the  line  of  air  discharge  is  vertically  down. 
All  intermediate  discharges  will  be  indicated  as  angular  discharge  as  follows: 
Either  top  or  bottom  angular  up  discharge  or  top  or  bottom  angular  down  discharge, 
the  smallest  angle  made  by  the  line  of  air  discharge  with  the  horizontal  being  specified. 

In  order  to  prevent  misunderstandings,  which  cause  delays  and  losses, 
the  arrangements  of  fan  drives  adopted  by  the  National  Association  of 
Fan  Manufacturers  and  indicated  in  Fig.  5  are  suggested. 

If  double  width,  double  inlet  fans  are  selected,  care  must  be  taken  that 
both  inlets  have  the  same  free  area.  If  one  inlet  of  a  forward  -curved  Made 
type  of  fan  is  obstructed  more  than  the  o|her,  the  fan -will  not  operate 
properly,  as  one  half  of  the^tgel^ill  delivet  more  air  than  the  other  half. 


, 

^Recommendations  adopted  by  the  National  Association  of  "Fan  Manufacturers. 

291 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


a 

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*l 

Arr.  4* 
For  direct  drive. 
overhung.  No  bearii 

mounted  on  motor 
Pedestal  for  motor 

-"  \ 

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Its 


Similar 
on  fan, 
couplinj 


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o 


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P3 

04 
< 

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292 


CHAPTER  17 — FANS  AND  MOTIVE  POWER 


The  backward  curved  and  double  curbed  types  with  backward  tip  operate 
satisfactorily  in  double  or  in  parallel  operation. 

MOTIVE  POWER 

It  is  no  easy  matter  to  predetermine  the  exact  resistance  to  be  encoun- 
tered by  a  fan  or,  having  determined  this  resistance,  to  insure  that  no 
changes  in  construction  or  operation  shall  ensue  which  may  increase  air 
resistance,  thus  requiring  more  fan  speed  and  power  to  deliver  the  required 
volume,  or  which  may  reduce  air  resistance,  thus  causing  delivery  of  more 
air  and  a  consequent  increase  of  power  even  at  constant  speed. 

It  is  recommended,  therefore,  for  centrifugal  type  fans  that  the  rated 
power  to  be  supplied  shall  exceed  the  rated  fan  power  by  a  liberal  margin , 
when  forward  cawed  types  are  used.  When  backward  or  double  curved 
blade  types  are  used,  motors  with  ratings  very  close  to  that  of  the  fan 
horsepower  demand  can  be  employed. 

Justification  for  liberal  power  provision  exists  also  in  the  possibility 
of  varying  demand  due  to  changes  in  ventilation  requirements,  intensity 
of  occupation,  and  weather  conditions. 

The  motive  power  of  fans  should  be  determined  in  accordance  with  the 
Standard  Test  Code  for  Disc  and  Propeller  Fans,  Centrifugal  Fans  and 
Blowers,  as  adopted  by  the  AMERICAN  SOCIETY  OF  HEATING  AND  VENTI- 
LATING ENGINEERS  and  the  National  Association  of  Fan  Manufacturers. 

Fans  may  be  driven  by  electric  motors,  steam  engines  (either  horizontal 
or  vertical),  gasoline  or  oil  engines,  and  turbines,  but  as  previously  stated 
the  drive  commonly  used  is  the  electric  motor. 

ELECTRIC  POWER 

Each  typje  of  electric  motor  and  kind  of  electric  current  has  its  advan- 
tages and  disadvantages  as  applied  to  a  fan.  For  motor  specifications  and 
standards,  the  Motor  and  Generator  Standards  of  the  National  Electrical 
Manufacturers  Association  should  be  consulted. 

Direct-connected  electric  motors  usually  are  very  efficient  for  fan 
driving  because  there  is  no  slippage  due  to  belts,  and  no  wear  or  noise  due 
to  chains  or  gears.  There  is  less  maintenance  and  upkeep  to  a  direct- 
connected  unit,  and  with  an  overhung  fan  wheel  on  the  motor  shaft,  the 
usual  fan  bearings  are  eliminated. 

The  disadvantage  of  a  slow-speed  direct-connected  motor  is  that  it 
may  be  unduly  large  and  heavy  as  well  as  costly,  but  this  may  be  offset 
by  the  compactness  of  the  unit  as  a  whole  due  to  limited  space  for  fan 
equipment. 

Should  anything  go  wrong  with  a  slow-speed  direct-connected  motor 
there  may  be  a  considerable  delay  in  securing  replacements,  as  these 
motors  are  not  usually  carried  in  stock,  as  is  the  case  with  moderately 
high-speed  motors. 

If  a  change  of  speed  is  found  necessary  with  a  direct-connected  motor, 
it  will  mean  a  change  of  motor,  which  may  necessitate  a  change  in  the 
motor  foundation  usually  built  with  the  fan  in  such  cases.  On*  the  other 
hand,  non-direct-connected  motors  have  transmissions  subject  to  wear 
and  slippage,  and  chains  or  gears  may  be  noisy  with  this  latter  type. 

293 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


TABLE  3.  CLASSIFICATION  OF  MOTORS 


GBOUP 

SUB- 
DIT. 

TYPE 

CUR- 
RENT 

SPEED 
CHAR- 

STARTING 
TORQUE 

STARTING 
CURRENT 

APPLICATIONS 

ACTERISTICS 

A 

1 

Shunt  wound 

d-c 

Constant 

Medium 

High 

Fans 

2 

Squirrel-cage 

a-c 

Constant 

Medium 

High  —  about 

Fans,  centrifu- 

six  times    full 

gal  pumps 

load 

3 

Synchronous 

a-c 

Constant 

Medium 

Starts  as  squir- 
rel cage  motor 

Motor    genera- 
tor    sets,     air 

compressors, 

fans 

4 

Slip     ring     or 

a-c 

Constant 

Heavy 

Low 

Vacuum  pumps, 

wound  rotor 

air  compres- 

sors 

5 

Double  squir- 

a-c 

Constant 

Heavy 

Medium 

Frequent  and 

rel-cage 

heavy    starting 

loads,     pumps, 

compressors 

6 

Low-torque 

a-c 

Constant 

Light 

Low 

Direct-con- 

capacitor 

nected  fans 

7 

High-torque 

a-c 

Constant 

Medium 

Low 

Belt    drive   of 

capacitor 

fans 

8 

High  -torque 

a-c 

Constant 

High 

Medium 

For    heavy 

capacitor 

starting     load 

such  as  larger 

fans,     pumps, 

compressors 

9 

Repulsion- 

a-c 

Constant 

High 

Medium 

Fans,    pumps, 

induction 

compressors 

B 

1 

Brush  shifting 

a-c 

Adjustable 

Medium 

Low 

Stokers,  boiler 

fans 

2 

Cumulative 

d-c 

Adjustable 

Heavy 

•    High 

Pumps 

comp'd     with 

shunt 

predominance 

3 

Squirrel  -cage, 
poles    can   be 

a-c 

Multi- 
speed 

Medium 

High 

Fans,  ice  ma- 
chines 

regrouped 

C 

1 

Series 

d-c 

Variable 

Heavy 

Low 

Fans 

2 

Cumulative 

d-c 

Variable 

Heavy 

Low 

Single-acting 

comp'd     with 

reciprocating 

series 

pumps 

predominance 

3 

Slip    ring  — 

a-c 

Variable 

Heavy 

Low 

Fans 

using  external 

resistance     in 

'  secondary 

294 


CHAPTER  17 — FANS  AND  MOTIVE  POWER 


However,  should  a  change  in  speed  be  necessary  where  the  motor  is  not 
direct-connected,  changes  in  speed  ratio  can  easily  be  accomplished  by 
changing  pulleys,  sprockets  or  gears  on  either  the  fan  or  the  motor.  In 
the  case  of  a  motor  breakdown  a  standard  stock  motor  may  easily  be 
substituted. 

A  type  of  drive  using  a  wedge-shaped  rope-like  belt,  singly  or  in  multi- 
ple, and  capable  of  use  on  short  pulley-centers  is  very  popular,  as  it 
enables  the  use  of  high  speed  motors  with  slow  speed  fans.  The  com- 
pactness secured  by  this  equipment  compares  favorably  with  that  of  a 
direct  connected  layout.  This  type  of  drive  also  is  very  quiet  in  operation, 
being  similar  to  a  conventional  belt  drive  in  this  respect.  Alternating 
current  motor  designs  are  such  that  improved  operating  characteristics 
are  obtained  with  the  higher  motor  speeds.  Efficiencies  and  power 
factors  are  improved  over  those  in  effect  with  slower  speed  motors,  thus 
showing  a  considerable  saving  in  power  consumption,  and  militating  in 
favor  of  some  effective  speed-reducing  transmission  device  such  as  is 
given  by  multiple  wedge-shaped  belts. 

Quietness  of  operation  is  more  readily  obtained  with  moderately  high 
speed  induction  motors  than  with  low  speed  motors,  as  any  slight  magnetic 
unbalance  in  the  latter  is  not  as  easily  heard.  Amplifications  of  motor 
induction  noises  in  parts  of  a  building  remote  from  the  motor  equipment 
sometimes  are  carried  by  the  steel  work,  ducts,  or  piping  in  the  building. 
There  is  considerable  evidence  that  these  sounds  are  more  easily  con- 
trolled with  high  motor  speeds  than  with  low  ones. 

Motors  which  are  practically  quiet  in  operation  and  free  from  magnetic 
disturbing  noises  can  be  obtained,  and  should  always  be  specified  for 
quietness  of  operation  when  used  for  fan  installations  in  buildings  where 
quietness  is  a  factor. 

In  the  construction  of  fan  and  motor  foundations  where  the  machinery 
is  mounted  on  the  floor  or  upon  a  concrete  platform,  it  is  a  usual  practice 
to  install  a  layer  of  cork  on  top  of  which  is  laid  or  floated  the  base  which 
carries  the  apparatus.  It  is  essential  that  the  bolts  or  lag  screws  which 
fasten  the  machines  to  this  foundation  shall  not  extend  through  to  the 
floor.  It  is  wise  to  fasten  curbs  to  the  floor,  these  presenting  insulated 
surfaces  to  the  machinery  foundation  and  so  preventing  it  from  traveling. 
Rubber,  especially  in  shear  or  in  tension,  is  valuable  as  a  sound  absorber 
in  foundations  for  machinery.  Steel  shoes  for  fans  and  motors  with 
rubber  inserts  are  available.  Steel  springs  are  also  used  effectively  for 
this  purpose. 

The  general  classification  of  motors  used  for  heating,  ventilation  and 
air  conditioning  is  shown  in  Table  3. 

Control  for  Electric  Motors 

Very  small  direct  current  motors  may  be  started  by  throwing  them 
directly  on  the  line  through  a  suitable  starting  switch.  The  larger  sizes 
require  some  type  of  starting  rheostat.  When  speed  adjustment  is 
desired,  the  controller  for  adjusting  the  speeds  of  the  motor  usually 
functions  also  as  a  starting  device. 

Alternating  current  motors  of  5  hp  and  under  usually  may  be  thrown 
directly  on  the  line.  It  is  good  practice  to  use  a  starting  switch  equipped 

285 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

with  a  thermal  overload  or  inverse  time  limit  overload  device.  This  type 
of  switch  provides  protection  to  the  motor  beyond  that  given  by  fuses. 
Fuses,  when  used,  necessarily  must  be  large  enough  to  take  care  of  the 
inrush  current  but  this  makes  them  inadequate  for  protecting  the  motor 
under  operating  conditions.  The  thermal  overload  device  allows  for  this 
inrush  and  does  not  function  until  an  overload  has  become  persistent,  the 
time  element  depending  upon  the  percentage  of  overload  beyond  the 
rating  of  the  element.  This  type  of  switch  is  available  for  manual  opera- 
tion and  also  is  furnished  in  the  magnetic  type  for  remote  operation  by 
push  button,  or  for  operation  by  other  types  of  pilots,  such  as  pressure 
switches  and  thermostats. 

On  standard  squirrel  cage  motors  above  5  hp  a  starting  compensator 
usually  is  employed  to  keep  the  inrush  current  within  the  limits  specified 
by  the  local  power  companies.  Compensators  may  be  obtained  in  trans- 
former types  and  primary  resistor  types,  and  usually  are  furnished  for 
manual  operation.  They  can  be  secured  for  remote  control  also,  but 
necessarily  are  expensive.  However,  the  new  type  of  high  reactance,  self- 
starting  motors  usually  may  be  thrown  across  the  line  up  to  30  hp  in  size, 
and  still  have  their  inrush  current  within  the  limits  of  the  rules  of  the 
National  Electric  Light  Association.  With  this  type  of  motor  a  magnetic 
contactor  usually  is  used.  This  device  may  be  operated  from  a  remote 
point  by  push  button,  if  desired.  These  magnetic  contactors  are  furnished 
usually  with  thermal  overload  and  no-voltage  protection. 

For  remote  operation  of  motors  through  magnetic  starters,  the  operat- 
ing buttons  may  be  located  in  the  engineer's  or  manager's  office,  and 
tell-tale  indicating  lamps  may  be  wired  up  with  the  circuit  to  indicate 
whether  or  not  the  unit  is  in  operation.  This  type  of  control  is  very 
desirable  in  large  buildings  where  the  engineer  is  to  .have  complete  charge 
of  the  ventilating  system. 

Remote  or  automatic  control  of  the  units  may  be  effected  also  by 
pneumatic  or  hydraulic  apparatus,  or  by  thermostats  or  by  pressure 
devices  which  are  provided  with  electric  contacts  for  starting  or  stopping 
the  units  upon  reaching  certain  conditions. 

Variable  speed  slip  ring  motors  and  direct  current  motors  may  also 
be  arranged  for  remote  speed  control  by  means  of  pre-set  automatic 
regulators,  where  the  operating  speed  of  the  motor  is  set  by  a  dial-switch 
(which  may  be  near  the  fan  or  at  a  remote  point)  and  the  motor  is  then 
automatically  controlled  at  any  given  speed  merely  by  operating  the 
remote  control  push  button  for  starting  or  stopping  the  equipment. 

Arrangements  may  be  made  for  remote  control  of  fan  motors,  or  for 
automatic  control  by  influence  of  temperature.  Remote  control  may  be 
by  pneumatic  or  by  hydraulic  manipulation  as  well  as  by  electrical  means. 

In  many  large  ventilating  systems  which  have  heating  plants  in  con- 
nection, steam  engines  are  used  to  operate  fans.  A  medium  speed  steam 
engine,  exhausting  at  low  pressure  into  the  heating  system  is  a  very 
economical  source  of  power,  is  quiet  in  operation,  and  has  a  wide  range  of 
speed  variation.  The  steam,  economy  of  such  an  engine  usually  is  of  little 
importance,  since  the  engine  serves  as  an  auxiliary  to  the  pressure- 
reducing  valve  interposed  in  such  cases  between  the  boiler  and  the  heaters. 

Internal  combustion  engines  and  line  shafting  often  are  used  for  fan 

296 


CHAPTER  17 — FANS  AND  MOTIVE  POWER 


driving,  requiring  clutches  or  shift-belts  with  loose  pulleys  in  order  to 
secure  proper  starting  and  control. 

Ability  to  adjust  the  speed  of  ventilating  fans  is  desirable  as  a  measure 
of  economy  and  adaptability  to  varying  loads,  but  where  such  adjust- 
ments are  provided  very  definite  speed  and  pressure  indications  should  be 
supplied  at  the  controller,  since  without  them  in  most  cases  the  operator 
would  be  compelled  to  guess  at  the  output. 

REFERENCES 

Heating,  Ventilating  and  Air  Conditioning,  by  Harding  and  Willard,  Revised  Edition, 
1932. 

Fan  Engineering,  Buffalo  Forge  Company. 

Theories  and  Practices  of  Centrifugal  Ventilating  Machines,  by  D.  Murgue,  trans- 
lated by  A.  L.  Stevenson. 

Mechanical  Engineers'  Handbook,  by  Kent. 

Mechanical  Engineers'  Handbook,  by  Lionel  S.  Marks. 

Constructive  Mechanism  and  the  Centrifugal  Fan,  by  George  D.  Beals. 

Coal  Miners  Pocket  Book. 

The  Fan,  by  Charles  H.  Innes. 

Mine  Ventilation,  by  J.  J.  Walsh  (A.S.H.V.E.  TRANSACTIONS,  Vol.  23,  1917). 

Fan  Blower  Design,  by  H.  F.  Hagen  (A.S.H.V.E.  TRANSACTIONS,  Vol.  28,  1922). 

The  Centrifugal  Fan,  by  Frank  L.  Busey. 

Section  X,  A.S.H.V.E.  Code  of  Minimum  Requirements  for  the  Heating  and  Venti- 
lation of  Buildings  (Edition  of  1929). 

PROBLEMS  IN  PRACTICE 

1  •  In  a  public  building,  what  type  of  fan  is  suitable  for: 

a.  A  supply  fan? 

b.  An  exbaust  fan? 

a.  The  centrifugal  housed  fan  is  well  suited  for  this  work.    The  various  types  are  the 
forward  curved  blade,  the  radial  blade,  the  full  backward  curved  blade,  and  the  medium 
speed  double  curved  blade  with  backward  tip.    When  direct  connected  motors  are  to  be 
used,  the  backward  tip  fans,  on  account  of  their  speeds,  are  better  adapted.    This  type 
has  the  added  advantage  of  having  a  limiting  horsepower  characteristic  which  will 
prevent  an  overload  on  the  motor.    Where  the  belt  drive  is  used,  all  of  the  above  types 
are  suitable. 

b.  For  exhaust  work  all  of  the  above  types,  as  well  as  disc  and  propeller  fans  are  suitable, 
although  the  latter  are  seldom  used  except  where  there  is  little  or  no  duct  work  con- 
nected to  the  fan. 

2  •  In  selecting  fans  for  quiet  operation  in  public  buildings : 

a.  Should  the  outlet  velocity  of  the  fan  be  limited? 

b.  Should  the  tip  speed  of  the  fan  be  limited? 

a.  Because  all  commercial  fans  operating  at  pressures  suitable  for  this  class  of  work 
would  be  considered  noisy  if  the  fan  were  to  discharge  directly  into  the  room,  and 
because  the  duct  system  on  the  fan  discharge  is  depended  upon  to  absorb  a  reasonable 
amount  of  fan  noise,  it  is  desirable  to  have  a  moderate  run  of  duct  work  with  some  bends 
and  elbows  included  as  sound  deadeners.    Where  this  duct  is  of  necessity  very  short,  the 
outlet  velocity  must  be  kept  down  to  the  lower  limits  recommended  in  this  chapter  or 
else  an  efficient  sound  absorber  must  be  used.    The  experience  of  the  engineer  must  be 
his  guide  in  determining  the  allowable  outlet  velocity  in  each  individual  case. 

b.  Tip  speed  should  not  ordinarily  be  limited,  because  different  types  of  fan  blades  have 
entirely  different  allowable  tip  speeds  for  quiet  operation.    A  fan  having  a  backward 
blade  at  the  tip  can  run  at  much  higher  tip  speed  than  can  a  forward  curved  or  a  straight 
blade  fan,  with  the  same  degree  of  quietness. 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

3  •  Is  a  direct  connected  or  a  belted  fan  preferable  in  public  building  work? 

Where  space  is  at  a  premium,  direct  connection  is  best.  Next  in  space  economy  is  the 
short  V-belt  drive.  The  flat  belt  drive  fan  requires  the  greatest  floor  space.  In  this 
class  of  work,  pressures  are  usually  so  low  that  even  with  the  high  speed  fans  the  motor 
cost  is  greater  for  direct  connected  units  than  for  belt  drive  fans. 

4  •  a.  What  type  fans  are  used  in  industrial  work? 

b.  What  outlet  velocity  is  suitable? 

a.  All  of  the  centrifugal  types  are  suitable;  the  disc  and  propeller  types  are  suitable  for 
low  pressure  work,  or  they  are  often  used  as  exhausters. 

b.  The  outlet  velocities  on  fans  for  industrial  work  can  be  much  higher  than  can  those  in 
public  building  work,  where  quietness  is  essential.    Fans  should  be  selected  with  outlet 
velocities  as  recommended  in  this  chapter,  using  the  upper  limit  of  velocities. 

5  •  Are  direct  connected  or  belted  fans  preferred  in  industrial  work? 

In  industrial  applications,  fans  are  often  advantageously  direct  connected  to  motors. 
The  pressures  are  usually  high  enough  to  use  standard  motor  speeds.  The  high  speed 
types  of  fans  have  limiting  horsepower  characteristics  so  that  little  margin  in  power  must 
be  provided  in  the  driving  motor.  Belted  fans  may  be  used,  but  where  high  power  is 
required  a  special  arrangement  is  often  necessary  for  shaft  and  bearings  on  account  of  the 
weight  of  the  sheave  and  the  belt  pull. 

o  •  A  forward  curved  multiblade  fan  which  requires  5.4  bhp  is  delivering  22,800 
cfm  at  70  F  against  a  resistance  pressure  of  1  in.  of  water  at  an  outlet  velocity 
of  1440  fpm: 

a.  What  is  the  static  efficiency? 

b.  What  is  the  total  efficiency? 

a.  66.3  per  cent  (see  Equation  1). 

b.  74.5  per  cent  (see  Equation  2). 

7  •  If  the  above  fan  has  a  54-in.  diameter  wheel  and  operates  at  193  rpm, 
will  it  be  suitable  for  a  ventilating  installation  where  a  minimum  of  noise  is 
desirable? 

Yes.  The  tip  speed  will  be  2720  fpm  and  this,  together  with  the  1440  fpm  outlet  velocity, 
falls  within  the  limits  given  in  Table  1  for  1-in.  resistance  pressure. 

8  •  Assuming  that  a  7^4  hp  constant  speed,  high  reactance  type,  self -starting 
electric  motor  is  used  to  drive  the  above  fan,  what  electrical  starting  apparatus 
should  be  used  for  control  from  a  remote  point? 

An  across-the-Kne  type  magnetic  push  button  starter  with  indicating  lamps  to  show 
whether  or  not  the  unit  is  in  operation. 

9  •  What  objectionable  feature  is  inherent  in  the  ordinary  propeller  fan  when 
it  is  operating  at  high  resistance  pressures? 

It  must  operate  at  a  high  speed  with  consequent  noise. 

10  •  At  what  point  should  a  fan  be  selected  for  operation,  and  why? 

At  its  point  of  maximum  efficiency  because  the  cost  of  operation  and  the  noise  produced 
will  be  least. 


298 


Chapter  18 

SOUND  CONTROL 

Measurement  of  Noise,  Noise  in  Buildings,  Coefficients  of 
Absorption,  Insulation  of  Air-Borne  Sound,  Location  and 
Insulation  of  Equipment  Room,  Insulation  of  Machinery  and 
Solid-Borne  Vibration,  Control  of  Noise  Transmission  Through 
Ducts,  Effect  of  Humidity  upon  Acoustics 

THE  ventilating  and  air  conditioning  of  any  space  affect  its  acoustics 
and  become  apparent  when  consideration  is  given  to  the  require- 
ments for  good  hearing  in  any  architectural  interior.    The  requirements 
which  must  be  given  careful  study  are: 

1.  The  room,  should  be  free  from  noise,  whether  of  inside  or  outside  origin. 

2.  The  useful  sound,  whether  speech  or  music,  should  be  sufficiently  loud  (with 
reference  to  any  residual  noise)  to  be  heard  easily  and  distinctly. 

3.  The  useful  sound  should  be  distributed  uniformly  in  all  parts  of  the  room,  and  the 
sound  reaching  the  listeners  should  be  free  from  long-delayed  reflections  which  produce 
interference  or  echoes. 

4.  The  room  should  be  free  from  pronounced  resonant  tones  which  may  result  from 
either  volume  or  panel  resonance. 

5.  The  room  should  contain  sound-absorptive  materials  in  such  amounts,  and  of  such 
qualities,  as  will  provide  a  proper  balance  between  the  persistence  and  cessation  of  the 
articulated  components  of  sound,  that  is,  the  reverberation  in  the  room  should  be  long 
enough  to  sustain  harmony  and  impart  tonal  blending  to  music,  and  at  the  same  time  it 
must  be  short  enough  to  prevent  the  overlapping  and  confusing  of  the  separate  sounds 
of  speech. 

Obviously,  the  first  of  these  requirements  is  the  one  which  imposes 
restrictions  on  the  installation  of  air  conditioning  or  ventilating  equip- 
ment— the  equipment  noises  must  be  unobjectionable  in  occupied  rooms — 
although  the  fifth  requirement  is  not  entirely  independent  of  the  humidity 
and  temperature  of  the  air. 

LOUDNESS 

Loudness  is  the  sensation  of  sound  intensity.  When  it  is  said  that  one 
sound  is  louder  than  another  a  difference  in  intensity  level  is  implied. 
Two  identical  whistles  when  sounded  together  do  not  make  a  sound  twice 
as  loud  as  one.  It  may  take  ten  to  make  a  sound  20  per  cent  louder  than 
one*  It  has  been  found  that  loudness  bears  a  logarithmic  relationship  to 
intensity  of  sound.  On  this  basis  a  scale  of  loudness  has  been  built  and  a 
unit,  the  decibel  (db),  has  been  established.  This  scale  is  illustrated  in 
Fig.  1  which  shows  the  loudness  of  some  typical  noises.  The  formula  for 
relating  loudness  and  intensity  is: 

Xi  -  L*  -  10  log*  A  (1) 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

where 

L  =  Loudness  in  db;  /  =  Intensity. 

Thus  the  two  whistles  made  a  noise  10  logic  2  =  3  db  louder  than  one 
whistle  and  the  ten  whistles,  10  logic  10  =  10  db  louder  than  one.  It 
would  take  a  hundred  whistles  to  make  a  noise  20  db  louder  than  one  and 
a  thousand  to  make  a  noise  30  db  louder. 

MEASUREMENT  OF  NOISE 

Since  the  chief  acoustical  problem  in  the  ventilating  or  air  conditioning 
of  a  building  consists  of  reducing  equipment  noise,  it  is  necessary  to 
describe  methods  for  measuring  noise.  The  measurement  of  noise  is  a 
relatively  new  problem,  and  although  there  are  several  reliable  methods, 
there  are  as  yet  no  standardized  units,  scales,  or  instruments  for  measuring 
noise1.  However,  the  decibel  (db)  described  above  is  widely  used  in  this 
country  and  England  as  the  standard  unit  for  noise  or  sound  intensity — a 
unit  of  the  same  size,  but  called  a  phon,  is  used  in  Germany — and  the  zero 
level  of  the  scale  is  >a  barely  audible  sound.  Since  the  relation  between 
subjective  loudness  and  sound  intensity  is  dependent  upon  pitch,  it  is 
customary  to  refer  loudness  to  a  single  frequency.  A  1000-cycle  tone  is 
generally  accepted  as  the  reference  frequency,  that  is,  the  loudness  of  any 
sound  is  rated  in  terms  of  an  equally  loud  1000-cycle  tone.  Thus,  a  noise 
of  50  db  means  that  the  noise  would  be  judged  to  be  of  the  same  loudness 
as  a  1000-cycle  tone  which  is  50  db  above  the  normal  threshold  of  audi- 
bility for  the  1000-cycle  tone. 

As  the  frequencies  decrease  below  1000  cycles,  the  ear  becomes  less 
sensitive,  until  at  about  30  cycles  sounds  are  no  longer  audible  regardless 
of  their  intensity.  Similarly,  for  higher  frequencies,  the  limit  of  audi- 
bility is  reached  around  7000  cycles.  Thus,  at  frequencies  below  1000 
cycles,  sounds  of  the  same  loudness  must  have  a  greater  intensity  than  at 
1000  cycles.  This  is  particularly  fortunate,  as  otherwise  the  low  fre- 
quency sounds  would  mask  all  others. 

Noise  measurements  are  usually  made  by  one  of  three  methods.  The 
first  is  the  electrical  instrument  method,  which  uses  a  noise  meter  usually 
consisting  of  a  microphone,  an  amplifier,  and  a  galvanometer.  Where 
such  a  meter  is  to  measure  the  loudness  of  a  noise  without  regard  to  the 
frequency  distribution,  it  must  contain  a  weighted  network  which  elec- 
trically simulates  the  varying  sensitivity  of  response  of  the  ear  to  different 
frequencies.  Where  it  is  desired  to  analyze  the  character  of  the  sound, 
filters  which  shut  out  all  but  certain  bands  of  frequencies  are  used  with  the 
meter.  A  number  of  manufacturers  make  such  meters. 

The  second  method  consists  essentially  of  varying  the  intensity  of  an 
artificially  generated  sound  until  the  noise  generated  is  masked  by  the 
noise  being  measured.  Obviously,  this  method  is  subject  to  human  errors 
in  observation  to  which  the  instrumental  method  is  not,  but  in  the  hands  of 


*See  Proposed  Tentative  Standards  for  Noise  Measurement,  and  Proposed  American  Tentative  Standard 
Acoustical  Terminology  of  the  American  Standards  Association  Sectional  Committee  on  Acoustical  Measure- 
ments and  Terminology. 

Also  see  How  Sound  is  Controlled,  by  V.  O.  Knudsen  (Heating,  Piping  and  Air  Conditioning,  October, 
1931),  and  Acoustical  Problems  in  the  Heating  and  Ventilating  of  Buildings,  by  V.  O.  Knudsen  (A.S.H.V.E. 
TRANSACTIONS,  Vol.  37,  1931). 

300 


CHAPTER  18 — SOUND  CONTROL 


a  careful  observer  quite  satisfactory  results  may  be  obtained.  One 
instrument  used  is  the  audiometer,  which  consists  of  a  buzzer,  an  ear 
phone,  and  a  rheostat.  The  phone  is  held  a  fixed  distance  from  the  ear 
while  the  resistance  of  the  rheostat  is  varied  until  the  sound  of  the  buzzer, 
as  transmitted  electrically  to  the  phone,  can  no  longer  be  heard.  Audio- 
meters are  available  either  for  covering  all  frequencies,  as  in  the  noise 
meter,  or  for  covering  certain  frequency  bands  only. 

A  third  method  of  measuring  noise,  simple,  yet  sufficiently  accurate  for 
most  field  measurements,  employs  only  three  tuning  forks  and  a  stop 
watch.  Forks  having  frequencies  of  128,  512,  and  2048  are  recommended. 
The  forks  must  be  calibrated.  That  is,  it  is  necessary  to  know  for  each 
fork  (1)  the  initial  intensity,  in  number  of  decibels  above  its  threshold, 
immediately  after  it  has  received  a  standard  hit  or  excitation,  and  (2)  the 
damping  rate,  in  decibels  per  second.  These  calibrations  can  be  made  in 
any  well-equipped  acoustical  laboratory.  A  standard  hit  or  excitation  can 
be  imparted  to  the  fork  by  a  felt-covered  spring  hammer,  or  simply  by 
letting  the  fork  fall  from  a  vertical  position  through  an  arc  of  90  deg, 
hitting  a  suitable  pad  (such  as  soft  rubber  or  felt  for  the  128  and  512  forks 
and  hard  rubber  for  the  2048  fork).  The  average  512  steel  fork  will  have 
an  initial  intensity,  when  held  %  in-  from  the  ear  with  the  broad  side  of 
the  prong  facing  the  ear  canal,  of  about  80  db,  and  will  decay  at  a  rate  of 
about  1.0  db  per  second.  Such  a  fork  will  remain  audible  about  80  sec 
in  a  perfectly  quiet  place,  provided  the  listener  has  normal  hearing.  In 
the  presence  of  a  noise,  it  will  remain  audible  until  its  tone  is  just  masked, 
by  the  noise.  Thus,  if  a  512  fork,  having  an  initial  intensity  of  80  db  and 
a  damping  rate  of  1.0  db  per  second,  should  be  found  to  remain  audible 
35  sec  in  the  presence  of  a  certain  noise,  the  masking  effect  of  the  noise 
is  80  -  35,  or  45  db. 

Procedure 

The  method  of  measuring  any  noise  is  as  follows:  The  observer,  in  the 
presence  of  the  noise,  strikes  the  128  fork  a  standard  blow.  At  the  same 
instant  he  starts  a  stop  watch.  The  fork  is  then  held  in  front  of  the  ear 
canal,  and  moved  back  and  forth  slightly,  until  the  tone  of  the  fof!Ti§~JTist 
completely  masked  by  the  noise,  at  which  instant  the  watch  is  stopped. 
This  measurement  is  repeated  at  least  two  times.  The  average  time  is 
subtracted  from  the  time  the  128  fork  remains  audible  in  a  quiet  place. 
This  difference  multiplied  by  the  damping  rate  of  the  fork  gives  the  mask- 
ing effect  of  the  noise  at  128  cycles.  Similar  measurements  are  made  with 
the  512  and  2048  forks.  Measurements  of  this  type  give  a  satisfactory 
description  of  both  the  intensity  and  the  frequency  distribution  of  the 
noise.  The  average  masking  effect  of  the  noise  at  128,  512,  and  2048 
cycles  will  usually  be  about  5  to  10  db  less  than  the  reading  given  by  a 
noise  meter. 

NOISE  IN  BUILDINGS 

Measurements  of  the  intensity  of  speech,  music  and  noise  in  many 
buildings,  with  special  consideration  of  the  noise  produced  by  ventilating 
equipment,  have  given  the  results  indicated  by  Fig.  1.  The  equivalent 
loudness  of  sounds  in  buildings  varies  from  less  than  10  db  near  the 
outlet  of  an  air  duct  in  a  very  quiet  sound  studio  to  nearly  100  db  in  a 
noisy  boiler  factory.  It  will  be  noted  that  the  noise  from  the  ventilating 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

fan  in  a  certain  high  school  auditorium  was  nearly  as  loud  as  average 
speech  in  a  large  auditorium.  Such  an  amount  of  noise  is  devastating  to 
good  acoustics;  in  fact,  it  is  impossible  to  hear  speech  in  the  presence  of 
such  a  noise. 


db 


Average  Loudness  of  Music  m  Room- 
Conversation  in  a  Small  Room- 

Speech  m  a  Small  Auditorium  -»• 
Speech  in  a  Large  Auditorium 


100 


*-  Boiler  Factory 


90 


80 


*•  Ventilating  Room  for  Large  Hotel  ( Very  Noisy ) 


70 


60    -e- 


50 


40 


30 


20 


10 


-Electric  Power  Substation 


-Inside  of  Duct,  near  Large  Low  Speed  Fan 

-Equipment  Room  ( Average  Condition ; 

-Fan  Room  for  School  Building  ( Rather  Quiet ) 

-Guest  Room,  Large  Hotel  on  Noisy  Street 
{ Windows  Open ) 

..Near  Outlet  of  Ventilating  Duct  m  High  School 
Auditorium  (Very  Noisy,  no  "Filters"  in  Duct) 


-  Fan  Noise  in  Theater  ( Poor  Control  of  Noise ) 


-Fan  Noise  in  Theater  (  Proper  Control  of  Noise) 

_Near  Outlet  of  Ventilating  Duct  in  M.  G.  M 
~  Sound  Studio  (  Planned  Control  of  Noise ) 


FIG.  1.  CHART  SHOWING  THE  EQUIVALENT  LOUDNESS  (IN  DECIBELS)  OF  SPEECH,  Music, 
AND  A  NUMBER  OF  NOISES  INCIDENT  TO  THE  VENTILATING  OF  ~ 


•Acoustical  Problems  in  the  Heating  and  Ventilating  of  Buildings,  by  V.  O.  Knudsen  (A.S.H  V.E 
TRANSACTIONS,  Vol.  37,  1931). 


In  every  problem  of  noise  reduction  in  buildings  it  is  necessary  to  know 
how  much  noise  can  be  tolerated.  The  noise  levels  given  in  Table  1  may 
be  regarded  as  completely  inoffensive.  They  represent  what  might  be 
termed  ideal  conditions,  not  often  realized  in  existing  buildings.  How- 
ever, they  represent  conditions  which  can  be  attained  by  proper  control 
of  noise,  and  the  heating  and  ventilating  engineer  should  aim  to  provide 
the  degree  of  quiet  specified  in  the  table. 

In  considering  the  tolerable  room  noise  level  due  to  heating,  ventilating, 
or  air  conditioning  apparatus,  not  only  must  the  absolute  value  of  the 
noise  be  considered  but  also  its  relation  to  the  room  noise  level  without 
the  apparatus  running.  This  is  necessary  since  a  large  increase  of  noise 
subjects  the  apparatus  to  serious  criticism  even  though  the  level  may  be 
low.  It  must  also  be  borne  in  mind  that  the  noise  produced  by  the  ap- 

302 


CHAPTER  18 — SOUND  CONTROL 


paratus  is  additive  to  that  of  the  room  without  apparatus.  Thus  if  the 
two  are  equal,  when  combined  the  noise  level  will  be  3  db  higher.  For 
these  reasons  the  room  noise  caused  by  the  apparatus  should  not  exceed 
the  other  room  noise. 

Noise  Control 

Essential  to  the  design  of  a  satisfactory  system  are:  first,  a  knowl- 
edge of  the  nature  and  intensity  of  the  noise  generated  by  the  various 
parts  of  the  equipment;  second,  a  knowledge  of  how  to  vary  the  noise 
level  between  the  apparatus  and  the  conditioned  room  if  need  be; 
third  j  a  knowledge  of  the  acceptable  level  of  apparatus  noise  in  the  con- 
ditioned room.  Besides  these,  the  engineer  must  be  able  to  deal  with 
other  noises  which  might  enter  the  room  when  openings  are  made  into  it, 
such  as  cross  talk  between  rooms  connected  with  common  ducts,  and  noise 

TABLE  1.    ACCEPTABLE  NOISE  LEVELS 


Talking  picture  studios , 

Radio  broadcasting  studios , 

Hospitals _. 

Music  studios 

Apartments,  hotels,  homes,  small  private  offices 

Theaters,  churches,  auditoriums,  classrooms,  libraries _ 

Talking  picture  theaters,  small  clothing  stores 

General  offices 

Large  public  offices,  banking  rooms,  upper  stories  of  department 
stores,  restaurants,  barber  shops 


Grocery  stores,  drug  stores 

Accounting  and  typewriting  offices- 
Main  floor  of  department  stores 


6  to    8  db 

8  to  10  db 

8  to  12  db 

10  to  15  db 

10  to  20  db 

12  to  24  db 

15  to  25  db 

20  to  30  db 

25  to  35  db 
30  to  50  db 
35  to  45  dD 
40  to  50  db 


transmitted  to  portions  of  duct  systems  outside  the  conditioned  room  and 
thence  to  its  interior. 

The  problem  of  apparatus  noise  is  receiving  the  study  of  equipment 
manufacturers  who  are  aiming  at  both  noise  reduction  and  standardiza- 
tion. Some  manufacturers  now  have  noise  ratings  available  for  their 
equipment,  while  some  pass  each  unit  of  equipment  of  certain  types 
through  sound  tests  during  the  course  of  manufacture. 

The  problem  of  noise  reduction  from  apparatus  to  room  must  take  into 
consideration  and  treat  separately  the  three  modes  of  travel  of  noise  to  the 
room :  first,  from  the  apparatus  through  the  air  to  the  walls  of  the  room 
and  thence  to  its  interior;  second,  through  the  building  structure  to  the 
room;  third,  through  ducts  or  openings  to  the  room.  Because  the  noise 
entering  by  each  of  these  three  channels  is  susceptible  to  quantitative 
analysis,  solutions  are  available.  Along  with  the  transmission  of  sound 
through  the  building  structure,  the  engineer  must  also  consider  the 
transmission  of  vibration,  which  may  also  be  objectionable.  The  solution 
is  not  complete,  however,  until  the  effect  of  the  noise  entering  the  room  on 
the,  room  noise  level  is  determined. 

ROOM  NOISE  LEVEL,  COEFFICIENTS  OF  ABSORPTION 

One  of  the  most  effective  means  of  reducing  noises  in  ventilating  equip- 
ment is  accomplished  by  the  proper  covering  of  the  interior  walls  and 

303 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

ceiling  of  the  equipment  room,  or  the  inner  walls  of  the  ducts,  with  sound- 
absorptive  materials.  The  intensity  /  of  a  continuous  sound  in  a  room  is 

E       PS* 

I  =          or  --—  2 

a  a 

where 

E  —  the  rate  of  emission  of  the  noise  source  =  I1 5'.    (The  intensities  of  noises  entering 

the  room  times  the  areas  through  which  they  enter.) 

a  =  the  total  amount  of  absorption  supplied  by  the  boundaries  and  contents  of  the 
room. 

=  otiSi  +  <x«Sz  +  a353   -f ,   wJiere  Si,  52,  58f are  the  areas  of  the 

boundary  materials  for  the  room,  and  «i,  a>,  «3, are  the  corresponding  coefficients 

of  absorption. 

Hence,  by  increasing  tenfold  the  absorptivity  of  the  boundaries  of  a  room  it  is  possible 
to  reduce  tenfold  the  average  intensity  of  sound  in  the  room;  that  is,  the  intensity  level 
would  be  reduced  10  db. 

Thus  it  is  possible  to  compute  the  noise  level  in  the  room  if  the  intensity 
of  noises  entering  the  room  or  generated  in  it  are  known. 

It  will  be  seen  that  the  noise  intensity  reduction  is  dependent  upon  the 
amount  of  sound  absorption  in  the  room,  and  that  the  first  units  of  absorp- 
tion are  more  effective  than  succeeding  units.  In  general,  the  room  noise 
level  will  be  from  10  to  20  db  lower  than  the  air  inlet  or  outlet  noise 
intensity,  the  10  db  being  in  the  case  of  bare  rooms  having  large  venti- 
lating or  air  conditioning  openings  in  relation  to  their  size,  and  the  20  db 
in  the  case  of  rooms  having  large  amounts  of  absorptive  material  with 
small  openings.  In  some  cases,  the  noise  level  reduction  may  run  up  to  as 
much  as  30  db,  but  then  the  higher  sound  intensity  adjacent  to  the 
openings  tends  to  nullify  the  effects  of  the  extra  reduction.  Where  these 
openings  are  large,  the  local  effect  on  the  noise  intensity  extends  some 
distance  from  the  opening;  for  instance,  a  four-square-feet  opening  might 
have  a  local  effect  within  ten  feet,  while  a  one-half-square-foot  opening 
would  have  a  local  effect  within  only  five  feet. 

The  coefficients  of  sound-absorption  for  a  number  of  standard  absorp- 
tive materials  used,  or  suitable  for  use,  in  equipment  rooms  are  given  in 
Table  2.  Coefficients  are  given  for  frequencies  of  128,  512,  and  2048 
cycles.  Where  the  frequency  of  the  noise  is  not  known,  the  values  for 
512  or  128  cycles  are  usually  used. 

INSULATION  OF  AIR-BORNE  SOUND 

The  transmission  of  air-borne  sounds  through  rigid  partitions  is  accom- 
plished primarily  by  the  diaphragm-like  vibrations  of  the  partition.  The 
weight  per  square  foot  of  the  wall  is  the  determining  factor,  and  the 
insulation  value  of  a  wall,  in  terms  of  the  transmission  loss  in  decibels, 
is  proportional  to  the  logarithm  of  the  weight  per  square  foot.  Other 
factors,  such  as  size,  stiffness,  composition,  manner  of  mounting,  and  the 
use  of  multiple  structures  separated  by  air  spaces  or  flexible  connectors, 
contribute  to  the  effective  insulation.  If  the  coefficients  of  sound  trans- 
mission of  different  types  of  structures  and  tjhe  noise  intensity  in  the  space 
adjoining  a  room  are  known,  it  is  possible  to  calculate  the  noise  intensity 
in  a  room  by  the  use  of  formula  1  and  the  following  formula: 

Jl  =  /»T  (3) 

304 


CHAPTER  18 — SOUND  CONTROL 


TABLE  2.    COEFFICIENTS  OF  SOUND  ABSORPTION 


MATERIAL 


THICKNESS 


COEFFICIENTS  OP  SOUND  ABSOBPTION 


UNCHES) 

128 
Cycles 

512 
Cycles 

2048 
Cycles 

Acoustex  60,  spray  painted                                       1 

0.16 

0  51 

0.72 

Acousti-Celotex,  Single  B  % 

0.11 

0.45 

0.68 

Acousti-Celotex,  Triple  B                                         1% 

0.20 

0  75 

0.67 

Acoustic  Flexfelt—  

0.27 

0.56 

0.68 

Acoustone                                                                      1 

0  66 

0.69 

^Vkoustolith  plaster                                                        3^ 

0.21 

0  29 

0  37 

Akoustolith  A,  Tile  1 

0.14 

0.48 

0.83 

Brick  wall,  unpainted                                                18 

0.024 

0  031 

0  049 

Calicel           1 

0.23 

0.72 

0.71 

Corkoustic,  Type  C  1J^ 

0.08 

0.61 

0.64 

Glass          

0.035 

0.027 

0.020 

Insulite  Acoustile,  Type  44                                         1^ 

0.26 

0  50 

0.61 

Kalite,  with  three  coats  lacquer    ...                            M 

0.35 

0.43 

0.45 

Macoustic  Plaster,  stippled  to  depth  of  %  in  3^ 
Masonite     ...                             Jfg 

0.13 
0.18 

0.31 
0.32 

0.58 
0.33 

Plaster,  gypsum  on  hollow  tile.    „  

0.013 

0.020 

0.040 

Plaster,  gypsum,  scratch  and  brown  coats  on 
metal  lath  on  wood  studs 

0.020 

0.040 

0.058 

Plaster,  lime,  sand  finish,  on  metal  lath  % 
Poured  concrete,  unpainted 

0.038 
0.010 

0.060 
0.016 

0.043 
0.023 

Rockoustile  ....                                              1 

0.18 

0.57 

0.72 

Sabinite                                                                         /^ 

0.34 

0.49 

Sanacoustic  Tile                                                           1J£ 

0.19 

0.79 

0.74 

Stuccoustic  Plaster,  Type  XB       % 

0.29 

0.59 

0.72 

Transite  Tile                                                          i       1 

0.19 

0.81 

0.72 

Trutone  Tile                                                                 1% 

0.31 

0.57 

0.64 

Wood  sheathing,  pine  .    \         % 

0.098 

0.10 

0.082 

Wood,  varnished  .  1     

i 

0.05 

0.03 

0.03 

^Architectural  Acoustics,  by  V.  O.  Knudsen,  pp.  219,  220,  240-251. 


where 

711  ~  noise  intensity  in  space  adjacent  to  room. 
T  —  coefficient  of  sound  transmission. 

Coefficients  of  sound  transmission  for  some  common  walls  are  shown 
in  Table  3. 

Example  1.  Suppose  the  brick  wall  between  an  equipment  room  and  an  adjacent 
auditorium  has  an  area  of  200  sq  ft  and  a  coefficient  of  sound  of  0.00001  (see  Table  3) ; 
that  the  auditorium  contains  2000  sabines2  of  absorption ;  and  that  the  noise  level  in 
the  equipment  room  is  70  db  above  zero  level. 

rii 
70  -  0  =  10  logio  -y-  (from  Formula  1) 

Tfl 

*         =    1Q7 


~  =  107  X  0.00001  =  100  (from  Formula  3) 
lo 


100  X 


=  10  (from  Formula  2) 


2  A  sabine  is  1  sq  f t  of  totally  absorptive  surface. 


305 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

Room  loudness  =  10  logio  10  =  10  db 

If  the  sound  absorption  in  the  auditorium  had  been  as  small  as  200  sabines,  the  sound 
intensity  in  the  auditorium  would  have  been  10  times  as  great  and  the  noise  level  in  the 
auditorium  would  have  been  20  db. 

If  the  rest  of  the  auditorium  has  an  area  of  20,000  sq  ft  with  a  surrounding  noise 
intensity  of  50  db  (/"  =  105)  the  noise  level  due  to  all  of  the  noise  entering  through  the 
wall  would  be  found  as  follows: 


-^-  =  105  X  0.00001 


10  (Through  equipment  wall)  -j-  1  X 


20,000 
2000 


=  20 


Room  loudness  X  10  logio  20  =  13  db 

Now  suppose  that  there  is  also  a  duct  having  20  sq  ft  outlet  connecting  the  room  with 
apparatus  having  a  noise  level  of  70  db  (/"  =  IO7)  and  suppose  that  there  is  an  assumed 
ttenua  tion  in  the  duct  equivalent  to  a  transmission  factor  of  0.0002.  Then, 


IO7  X  0.0002  =  2  X  IO3 


-f-  »  20  (from  above)  +  2  X  IO3  X 


20 
2000 


40 


Room  loudness  =  10  logio  40  =  16  db 

It  may  be  seen  how  the  energies  of  noises  entering  a  room  are  added  to  obtain  the 
final  room  noise  intensity. 

The  average  coefficients  of  sound  transmission  (128  to  4096  cycles)  for 
a  number  of  walls  and  of  floor  and  ceiling  partitions  are  listed  in  Table  3. 


TABLE  3.  AVERAGE  COEFFICIENTS  OF  SOUND  TRANSMISSION  FOR  BUILDING  PARTITIONS* 


DESCRIPTION  OF  PARTITION 


AVERAGE 
COKFFICIBNT 


Brick  panel,  Mississippi,  8  in.;  plastered  both  sides  gypsum  brown  coat, 

smooth  white^finish;  good  workmanship 

Brick  wall,  2j^-in.  plaster  both  sides 

Brick  wall,  2J^-in.,  2-in.  furring  strips,  J^-in.  rigid  insulation  lath  plastered 

both  sides „ 

Brick  wall,  4  in.,  2-in.  furring  strips  and  J4-in.  rigid  insulation  lath,  plaster, 

on  one  side;  other  side  plastered  directly  on  brick. _ 

Concrete  flat  slab  floor  construction,  reinforced;  floating  floor  consisting 

of  nailing  strips,  rough  and  finish  flooring;  J^-in.  rigid  insulation  furred 

out  and  applied  as  ceiling. 

Glass,  plate  ££-in. 

Glass,  plate  M-in.  double  glazed,  IJ^-in.  separation 

Metal  lath,  double,  on  IJ^-in.  channels,  M-in*  gypsum  plaster;  without 

cross  bracing  dips;  4  in.,  connected  at  edges  only 

Tile,  hollow  clay  partition,  three  cells,  4  in.  x  12  in.  x  12  in.,  wood  furring 

strips,  J^-in.  rigid  insulation,  gypsum  brown  coat,  smooth  white  finish 

Wood  joists,  lower  side  plastered  on  wood  lath;  floating  floor  consisting  of 

nailing  strips,  rough  and  finish  flooring. 

Wood  studs,  four-paper  plaster  board,  three-coat  smooth  finish  gypsum 

plaster „ 

Wood  studs,  two  $4-in.  sheets  rigid  insulation  both  sides,  joints  filled, 

gypsum  scratch  and  brown  coats,  smooth  white  finish 

Wood  studs,  2  in.  x  4  in.,  staggered,  metal  lath,  J^-in.  gypsum  plaster; 

7  J£  in. ;  connected  at  edges  only 


0.000010 
0.000032 

0.0000016 
0.0000040 

0.0000020 

0.0010 

0.0001 

0.000016 

0.0000050 

0.0000050 

0.000010 

0.000013 

0.000040 


*Archiieciu  al  Acoustics,  by  V.  O.  Knudsen,  pp.  308-322. 

306 


CHAPTER  18 — SOUND  CONTROL 


LOCATION  AND  INSULATION  OF  EQUIPMENT  ROOM 

The  equipment  room,  if  possible,  should  be  located  at  a  considerable 
distance  from  all  rooms  in  which  quiet  is  required.  If  this  is  not  possible, 
it  is  necessary  to  provide  a  high  degree  of  insulation  against  the  noise 
which  may  be  transmitted  through  the  walls  of  the  equipment  room,  and 
also  against  the  noise  which  almost  certainly  will  be  communicated 
through  the  short  ducts.  (See  discussion  of  Control  of  Noise  Trans- 
mission through  Ducts,  p.  311.)  Three  wall  sections  and  two  floor  and 
ceiling  sections  which  are  satisfactory  for  the  wall  insulation  of  the 
equipment  room  are  shown  in  Fig.  2.  Other  partitions,  with  their  sound 
insulating  values,  are  listed  in  Table  3.  The  addition  of  absorptive 
materials  (such  as  are  described  in  Table  2)  to  the  inner  walls  and  ceiling  of 
the  equipment  room  will  not  only  increase  the  insulation  through  the 
walls,  but  will  also  reduce  the  intensity  of  the  noise  in  the  room.  The 
equipment  room  noise  intensity  may  be  figured  in  the  same  way  as  that  of 
the  conditioned  space,  taking  the  equipment  as  the  source  of  noise.  In 
case  the  equipment  is  subject  to  considerable  vibration  it  is  advisable  to 
provide  a  separate  or  floated  floor. 


-  4"  Brick 


Plaster 


insulation  Value =47  db. 


-      4"  Hollow  Clay  Tile 
1"* 


L"*  2"  Furnng  Strips 
s  Paper  and  Metal  Lath 
^Piaster 
Insulation  Value  =  52  db. 


-Absorptive  Blanket 
2  Fibre  Board 

S?"  Plaster 
\  2 

x  Staggered  Wood 
Studs 


Insulation  Value 
Greater  than  50  db. 


Finish  Ftoonng 
Absorptive  Blanket 
Plaster  on  Lath 


Insulation  Value  =  50  db. 


Flooring 


Resilient  Chairs 
x  Concrete  Slab 
Resilient  Hangers 
'Plaster  on  Lath 


Insulation  Value  -  60  db.,  or  more 


FIG.  2.    THREE  WALL  SECTIONS  AND  Two  FLOOR  AND  CEILING  SECTIONS  WHICH  ARE 
SUITABLE  FOR  THE  INSULATION  OF  EQUIPMENT  ROOMS* 

aAcoustical  Problems  in  the  Heating  and  Ventilating  of  Buildings,  by  V.  O.  Knudsen  (A.S.H.V.E. 
TRANSACTIONS,  Vol.  37,  1931). 


INSULATION  OF  MACHINERY  AND  SOLID-BORNE  VIBRATION 

Since  mechanical  vibrations  are  readily  transmitted  through  the  solid 
structure  of  a  building,  it  is  extremely  important  in  air  conditioning  that 
all  mechanical  equipment  in  which  vibrations  are  generated  be  thoroughly 
insulated  from  the  solid  structure  of  the  building.  An  almost  universal 
notion  prevails  that  the  vibrations  generated  by  machinery  can  be  in- 
sulated from  a  building  simply  by  placing  a  slab  of  cork  or  a  layer  of 
hair  felt  between  the  machinery  and  the  floor  of  the  room.  If  the  machinery 
is  sufficiently  heavy,  and  the  cork  or  felt  sufficiently  resilient,  this  ex- 
pedient may  suffice.  On  the  other  hand,  if  the  machinery  is  not  suf- 

307 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

ficiently  heavy  to  load  the  cork  or  felt  support  to  the  extent  that  the 
natural  frequency  of  the  machinery  on  the  cork  or  felt  is  low  in  com- 
parison with  the  frequency  generated  by  the  equipment,  the  cork  or  felt 
may  be  of  little  avail.  The  insulation  of  vibration  can  be  accomplished 
by  means  of  suitable  elastic  supports  or  suspensions,  but  the  design  of 
these  elastic  supports  should  be  based  upon  calculation  rather  than 
guess-work. 

The  theory  of  the  insulation  of  vibration  was  first  worked  out  by 
Soderberg3.  If  a  machine  of  mass  m  be  supported  by  an  elastic  pad  the 
amount  of  vibratory  force  communicated  by  the  machine  to  the  floor  or 
foundation  upon  which  it  rests  will  be  determined  by  the  elastic  and  viscous 
properties  of  the  pad.  The  ratio  of  the  vibratory  force  communicated  to 
the  floor  or  foundation  with  the  machine  resting  upon  the  pad,  and  with 
the  machine  resting  directly  upon  the  floor,  is  given  by  the  following 
equation :  

/         r.   |     j    1^ 

f4) 


where 

•tr  =  the  so-called  transmissibiltty  of  the  support. 

c  =  the  compliance  (that  is,  the  reciprocal  of^the  force  constant). 

r  —  the  mechanical  resistance  owing  to  the  viscous  forces  within  the  support. 

n  —  the  frequency  of  vibration  generated  by  the  machine  which  is  to  be  insulated, 
such  as  the  commutation  frequency  of  a  motor  or  the  blade  frequency  of  a  fan. 

m  —  the  mass  of  the  machine  to  be  insulated. 

It  should  be  noted  that  not  only  must  vibrations  within  the  audible  range  of  fre- 
quencies be  considered,  but  those  in  the  sub-audible  range  as  well,  since  these  may  cause 
objectionable  vibrations.  All  the  possible  frequencies  should  be  considered  in  the  calcu- 
lation. Sometimes  beat  effects  are  introduced  by  slight  irregularities  of  belts  or  pulleys 
that  have  much  lower  frequencies  than  those  of  the  rotating  elements. 

If  the  pad  is  to  be  of  any  value  in  the  prevention  of  solid-borne  vibra- 
tions, the  value  of  T!  must  be  considerably  smaller  than  unity.  If  the 
fundamental  frequency  of  vibration  generated  by  the  machine  happens  to 
coincide  with  the  natural  frequency  of  the  mass  of  the  machine  resting  on 
the  elastic  pad,  a  condition  of  resonance  will  be  established,  and  the 
machine  will  exert  a  greater  force  upon  the  foundation  than  it  would  if 
the  pad  were  completely  removed.  It  is  necessary,  therefore,  that  the 
elastic  support  be  sufficiently  compliant,  and  the  mass  of  the  machine 
sufficiently  heavy,  that  the  natural  frequency  of  the  mass  m  upon  its 
elastic  support  will  be  low  in  comparison  with  the  frequencies  which  are 
generated  by  the  machine.  Thus,  if  the  principal  vibrations  in  the 
machine  be  of  the  order  of  100  vibrations  per  second,  the  natural  frequency 
of  the  machine  mounted  on  its  elastic  support  should  not  exceed  about 
20  vibrations  per  second. 

If  a  slab  of  insulating  material  be  placed  under  the  entire  foundation  of 
a  machine,  as  is  often  done  in  practice,  it  may  happen  that  the  natural 
frequency  of  the  machine  on  its  elastic  support  will  be  nearly  the  same  as 
the  frequencies  which  are  to  be  insulated,  in  which  case  the  elastic  support 


»C.  R.  Soderberg,  The  Electric  Journal  (January,  1924),  and  succeeding  articles.  See  also  V.  O.  Knudsen, 
Physical  Review,  Vol.  32,  1928,  p.  324,  and  A.  L.  Kimball,  Journal  Acoustical  Society  of  America,  Vol  2, 
1930.  p.  297. 

308 


CHAPTER  18 — SOUND  CONTROL 


will  be  worse  than  nothing.  In  general,  as  Equation  4  shows,  both  m  and 
c  should  be  as  large  as  possible  if  the  vibrations  of  the  machine  are  to  be 
effectively  insulated  from  the  solid  structure  of  the  building.  Further- 
more, the  machine  should  rest  upon  a  rigid  floor  so  that  the  elastic 
yielding  of  the  floor  is  prevented  from  communicating  the  machinery 
vibrations  to  the  solid  structure  of  the  building. 

The  elastic  support  under  the  machine  acts  as  a  low-pass  filter  which 
passes  all  frequencies  below  about  two  times  the  natural  frequency  of  the 
machine  mounted  on  its  elastic  support,  but  prevents  all  frequencies 

above  about  V  .???  from  reaching  the  solid  structure  of  the  building.   The 

principal  influence  of  the  internal  mechanical  resistance  r  is  to  limit  the 
vibration  at  the  resonant  frequency.  It  is  generally  advisable,  therefore, 
to  use  materials  which  have  an  appreciable  internal  resistance. 

The  values  of  c  and  r  can  be  determined  for  any  specimen  of  flexible 
material  and,  when  known,  can  be  used  to  determine  the  insulation  value 
of  any  particular  set-up.  The  value  of  c  can  be  obtained  by  making  static 
measurements  of  the  amount  of  displacement  of  the  compressed  support 
for  each  additional  unit  of  the  compressing  force.  If  this  be  done  for  a 
specimen  of  the  flexible  "material  of  a  certain  thickness  and  area  of  cross 
section,  the  compliance  can  be  determined  for  any  other  thickness  or  area 
from  the  relation  that  c  will  be  directly  proportional  to  the  thickness  and 
inversely  proportional  to  the  area  of  the  flexible  support.  When  the 
internal  resistance  r  is  not  too  large,  it  can  be  determined  by  observing  the 
successive  amplitudes  of  the  free  vibrations  of  a  mass  m  which  rests  upon 
a  specimen  of  the  flexible  material,  and  solving  for  r  by  the  usual  log- 
decrement  method.  Or,  if  the  damping  be  so  great  that  the  free  motion  of 
m  is  non-oscillatory,  r  can  be  obtained  from  measurements  on  the  experi- 
mentally-determined resonance  curve  of  the  forced  vibrations  of  m,  or 
from  measurements  of  the  rate  of  return  of  m  when  it  is  given  an  initial 
displacement. 

If  the  resistance  of  a  certain  specimen  of  material,  as  cork,  felt,  or 
rubber,  has  been  determined  by  any  of  these  methods,  the  resistance  for 
any  other  thickness  or  area  of  the  material  can  be  determined  approxi- 
mately because  the  resistance  will  be  inversely  proportional  to  the 
thickness  and  directly  proportional  to  the  area  of  cross-section  of  the 
flexible  support.  Thus,  if  the  values  of  c  and  r  for  a  flexible  material 
be  known,  it  is  possible  to  calculate,  by  means  of  Equation  4,  the  amount 
of  insulation  that  will  be  obtained  from  the  use  of  this  material  as  a 
flexible  support  for  a  piece  of  equipment  having  a  mass  m.  For  the 
routine  calculations  in  practice,  r  may  be  neglected  with  only  a  slight 
sacrifice  of  accuracy.  Table  4  gives  the  values  of  c  and  r  for  a  number  of 
commonly  used  flexible  materials. 

In  general,  there  are  two  principal  points  to  observe  in  the  design  of  a 
flexible  support  for  any  piece  of  equipment,  namely,  the  material  should 
have  a  relatively  large  compliance  and  it  should  be  loaded  to  nearly  the 
upper  safe  limit  of  loading.  Several  flexible  metallic  supports  have  recently 
been  developed. 

Example  2.  A  machine  weighing  1000  Ib  has  a  base  area  of  20  sq  ft.  Assume  that  the 
principal  vibration  of  the  machine  has  a  frequency  of  100  cycles  per  second  (most 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

TABLE  4.    COMPLIANCE  AND  RESISTANCE  DATA  FOR  TYPICAL  SPECIMENS  OF 
FLEXIBLE  MATERiALSa 

The  compliances  and  resistances  given  in  the  table  are  for  specimens  1  in.  thick 
and  1  sq  cm  in  cross-section 


MATERIAL 

DESCRIPTION 
OF  MATERIAL 

APPROXIMATE  UPPEB 
SAFE  LOADING  IN 
POUNDS  PER  SQUARE 
INCH 

COMPLIANCE  c   IN 
CENTIMETERS  PER 
DYNE 

RESISTANCE    r   IN 
ABSOLUTE  UNITS 

Corkboard 
Corkboard 
Flax-li-num 
Celotex 

l.lOlbper 
board  foot 
O.TOlbper 
board  foot 
1.351bper 
board  foot 
Carpet  lining 

12 
8 
4  to  6 
10 

0.25x  10~6 
0.50x  10~6 
0.60x10-' 
0.40  x  10~6 

O.lox  105 
0.25x  105 
O.oOx  10* 

Celotex 

Insulating 

12 

0.18  x  10~6 

Insulite 

board 
Insulating 

15 

0.16x  10-* 

Mason  it  e 

board 
Insulating 

15 

0.12x  10-« 

Anti-Vibro-Block 
Sponge  Rubber 

board 
"25~l"b"per" 

5 
1  to  3 

0.60x  10~6 
3.6    x  10~6 

1.5  x  105 

Soft  India  Rubber 

cubic  foot 
55  Ib  per 

3  to  6 

1.2    x  10~6 

Hairfelt 

cubic  foot 
10  Ib  per 

1  to  2 

1.5    x  10~6 

cubic  foot 

^Architectural  Acoustics,  by  V.  O.  Knudsen,  p.  278. 


machinery  vibrations  are  less  than  150  vibrations  per  second,  and  the  assumed  frequency 
of  100  is  quite  representative  of  typical  machines).  Suppose  that  a  1-in.  slab  of  cork- 
board  weighing  1.10  Ib  per  board  foot  be  placed  between  the  machine  and  the  floor. 
The  loading  on  the  cork  will  then  be  only  50  Ib  per  square  foot,  or  slightly  more  than 
%  Ib  per  square  inch.  (It  is  assumed  that  the  compliance  c  in  centimeters  per  dyne  for  a 
specimen  1  in.  thick  and  1  sq  cm  in  cross-section  is  0.25  X  10~6  and  the  resistance  r  in 
mechanical  ohms  is  0.15  X  10s.) 

The  transmissibility  is  calculated  in  the  following  manner: 

Mass  of  machine  in  grams  =  1000  X  454  =  4.54  X  10&. 
Area  of  base  in  square  centimeters  =  20  X  144  X 
2.54  X  2.54  =  1.86  X  104. 

Therefore,  the  compliance  of  the  entire  support,  1  in.  thick  and   20  sq  ft  in  cross 

section,  is  0.25  X  lO"6  X  -T-^-TT-T^T  =  0-134  X  lO"10  cm  per  dyne,  and  the  resistance  of 

l.&o  X  lu* 

the  entire  support  is  0.15  X  10fi  X  1.86  X  104  =  0.28  X  109  mechanical  ohms  (or  absolute 
units).  Therefore, 


V 


<°-28  X 


+ 


X  100  X  0.134 


(0.28  +  109)2  4-  (2x  X  100  X  4.54  X  106  - 


=  0.93 


2-rc  X  100  X  0.134 

Consequently,  it  is  seen  that  the  transmissibility  is  nearly  equal  to  unity,  and  that  the 
support  therefore  is  not  satisfactory  for  insulating  100  or  fewer  vibrations  per  second. 

If  the  amount  of  cork  be  reduced  so  that  it  is  loaded  to  10  Ib  per  square  inch,  the  total 
area  of  the  supporting  cork  will  be  only  100  sq  in.  or  645  sq  cm.    The  compliance  of  the 

310 


CHAPTER  18— SOUND  CONTROL 


entire  support  will  now  be  0.25  X  lO"6  X  ^  »  0.39   X   10~fl  cm  per  dyne,  and  the 

resistance  will  be  0.15  X  105  X  645  -  0.97  X  107  mechanical  ohms  (or  absolute  units). 
Therefore 


-v 


(0.97  X  107)2  -f  10 


X  100  X  0.39 


(0.97  X  107)*  +  (  2x  X  100  X  4.54  X  10s  -  10* 


X  100  X  0.39  / 

It  is  seen,  therefore,  that  with  the  bearing  surface  on  the  cork  reduced 
to  100  sq  in.  (that  is,  with  the  cork  loaded  to  10  Ib  per  square  inch),  the 
transmissibility  is  reduced  to  0.037,  or  the  amplitude  of  vibration  trans- 
mitted to  the  floor  will  be  only  about  1/27  of  what  it  would  be  if  the 
machine  were  mounted  directly  upon  the  floor.  These  two  numerical 
examples  will  serve  to  show  not  only  the  manner  of  making  the  calcu- 
lations, but  also  the  importance  of  selecting  the  proper  type  and  design  of 
flexible  supports  for  insulating  the  vibrations  of  a  machine  from  the 
rigid  structure  of  a  building. 

CONTROL  OF  NOISE  TRANSMISSION  THROUGH  DUCTS 

The  most  troublesome  sources  of  noise  from  ventilating  and  air  con- 
ditioning equipment  are  fan  and  motor  noises  which  are  transmitted 
through  the  ducts.  The  reduction,  in  decibels,  of  noise  transmitted 
through  a  duct,  neglecting  reflection  from  ends  and  bends,  is  proportional 
(1)  directly  to  the  length  of  the  duct,  (2)  directly  to  the  perimeter  of  the 
duct,  (3)  inversely  to  the  area  of  cross-section  of  the  duct,  and  (4)  directly 
(or  at  least  approximately  so)  to  the  coefficient  of  sound  absorption  of  the 
material  which  comprises  the  interior  surface  of  the  duct.  It  is  apparent 
therefore  that  long  narrow  ducts,  lined  with  highly  absorptive  material, 
will  provide  a  high  degree  of  insulation  against  the  transmission  of  noise 
through  ducts.  In  fact,  small  ducts  (4  in.  x  6  in.),  made  of  material 
having  a  coefficient  of  sound-absorption  of  0.50,  will  provide  a  noise 
reduction  of  slightly  more  than  1  db  per  linear  foot. 

As  can  be  seen  from  an  inspection  of  Table  2,  noises  of  low  frequency 
are  difficult  to  absorb;  on  the  other  hand,  these  frequencies  are  easily 
reflected  by  elbows,  branches,  and  duct  ends  whereas  higher  frequencies 
are  little  affected.  Furthermore,  the  reflection  effects  are  more  pro- 
nounced in  small  ducts  than  in  large  ducts.  Hence,  by  introducing  into 
a  duct  a  sufficient  length  of  small,  absorptive  channels  together  with  a 
number  of  elbows  or  other  reflecting  elements  it  is  possible  to  reduce  the 
transmitted  noise  to  any  required  degree.  This  applies  not  only  to  ducts 
between  the  equipment  room  and  other  rooms  in  a  building,  but  also  to 
ducts  connecting  adjacent  or  nearly  adjacent  rooms.  By  the  proper  use 
of  such  filters  it  is  possible  to  eliminate  all  of  the  difficulties  which  arise  in 
connection  with  the  transmission  of  sound  through  ventilating  ducts.  The 
problem  is  an  engineering  one  which  can  be  worked  out  prior  to  the  in- 
stalling of  the  equipment,  and  it  can  be  calculated  in  such  a  way  as  to 
meet  the  most  rigorous  demands  for  silent  operation.  There  is  a  need  for 
quantitative  data  regarding  the  attenuation  or  noise-reduction  provided 
by  different  types  of  ducts,  but  even  with  the  meager  data  available  it  is 

311 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

possible  to  design  filters  which  will  suppress  the  ordinary  noises  incident 
to  the  ventilating  or  air  conditioning  of  buildings4. 

In  general,  the  motion  of  air  resulting  from  the  ventilating  of  rooms  is 
not  sufficient  to  introduce  any  appreciable  difficulty  in  auditoriums,  except 
where  noise  may  originate  from  the  issuing  of  high-speed  air  from  nozzles. 
However,  by  proper  stream-lining  of  the  nozzles,  it  is  possible  to  work 
with  speeds  which  are  adequate  for  all  practical  purposes  without  pro- 
ducing any  disturbing  noises.  Since  sound  is  propagated  with  a  velocity 
of  more  than  1100  fps,  the  velocity  of  the  air  would  have  to  attain  speeds 
of  at  least  20  to  30  fps  before  these  wind  velocities  would  have  any 
appreciable  influence  upon  the  propagation  of  sound. 

If  there  is  to  be  any  appreciable  motion  of  air  in  an  auditorium,  it  is 
advantageous  to  have  the  upper  layers  of  air  moving  in  a  direction  from 
the  stage  toward  the  audience,  as  this  will  tend  to  refract  the  sound  waves 
down  toward  the  audience.  However,  unless  the  speed  of  the  air  is  as 
great  as  20  or  30  fps,  the  amount  of  refraction  will  not  be  noticeable. 
Therefore,  as  a  rule  the  motion  of  air  in  an  auditorium  does  not  have  an 
appreciable  effect  upon  the  acoustical  properties  of  the  room. 

EFFECT  OF  HUMIDITY  UPON  ACOUSTICS 

Recent  experiments5  have  shown  that  both  the  humidity  and  the  tem- 
perature of  air  have  a  marked  influence  upon  the  rate  of  absorption  of 
high-pitched  sounds.  Perfectly  dry  air  is  less  absorptive  than  air  con- 
taining any  amount  of  water  vapor.  At  relative  humidities  of  5  to  25  per 
cent,  the  air  is  highly  absorptive  but  becomes  less  and  less  absorptive  as 
the  humidity  is  increased.  High-frequency  sounds  are  propagated 
better  in  cold  humid  air  than  in  hot  dry  air,  and  since  high-frequency 
sounds  are  particularly  important  for  the  preservation  of  good  quality 
in  speech  and  music  it  is  advantageous  to  maintain  the  air  in  a  room  at  a 
relatively  high  humidity,  not  less  than  about  55  to  60  per  cent.  On  the 
other  hand,  where  it  is  desirable  to  absorb  all  frequency  components  of 
sound,  as  for  the  reduction  of  noise  in  offices,  it  is  advantageous  to  main- 
tain relatively  dry  air. 

The  time  of  reverberation  in  a  room  is  given  by  the  following  equation  : 

.  =  0.0497  , 


Sloged  -  a) 

where 

V  =  volume  of  room  in  cubic  feet. 

S  —  interior  surface  of  room. 

a  =  average  coefficient  of  sound-absorption  of  the  interior  surface  of  the  room. 

m  —  the  absorption  coefficient  of  the  air  in  the  room. 

The  coefficient  m  depends  upon  the  frequency  of  the  sound  and  the 
humidity  (and  probably  the  temperature)  of  the  air.  At  a  temperature  of 
70  F,  and  for  sound  waves  having  a  frequency  of  4096  vibrations  per 
second,  m  =  0.0027  at  25  per  cent  relative  humidity,  0.0018  at  54  per 


*How  Sound  is  Controlled,  by  V.  O.  Knudsen  (A.S.H.V.E.  TRANSACTIONS,  Vol.  37,  1931). 
^Effect  of  Humidity  upon  the  Absorption  of  Sound  in  a  Room,  by  V.  O.  Knudsen  (Journal  Acoustical 
Society  of  America,  July,  1931).    Also  see  report  presented  at  the  May,  1933,  meeting  of  A.  S.  of  A. 

312 


CHAPTER  18 — SOUND  CONTROL 


cent,  and  0.0013  at  82  per  cent.  It  will  be  seen,  therefore,  that  the  absorp- 
tion of  sound  in  the  air  is  twice  as  great  at  a  relative  humidity  of  25  per 
cent  as  it  is  at  a  relative  humidity  of  82  per  cent.  This  explains  why 
sounds  in  the  open  travel  so  much  better  on  humid  days  than  they  do  on 
dry  days.  Although  this  dependence  of  absorption  upon  humidity  is 
characteristic  of  low-frequency  as  well  as  high-frequency  sound,  the  actual 
amount  of  absorption  in  the  air  is  negligible  for  frequencies  below  about 
1024  vibrations  per  second.  However,  the  absorption  of  the  higher 
frequencies  in  the  air  is  a  significant  factor,  and  its  dependence  upon 
humidity  calls  for  careful  consideration  in  planning  the  air-conditioning 
equipment  for  buildings. 


PROBLEMS  IN  PRACTICE 

1  •  What  are  the  requirements  for  good  hearing  in  a  room? 

Freedom  from  noise,  adequate  loudness  of  speech  or  music,  uniform  distribution  of 
sound  throughout  the  room,  freedom  from  echoes  and  sound  foci,  no  pronounced  reso- 
nance, and  proper  reduction  of  reverberation. 

2  •  Why  do  modern  improvements  in  the  acoustics  and  air  conditioning  of 
buildings  present  new  acoustical  problems   to  the  heating  and  ventilating 
engineer? 

In  acoustically  treated  rooms,  both  outside  and  inside  noise  are  reduced,  and  conse- 
quently the  noise  of  ventilating  equipment  becomes  more  noticeable.  The  closed 
windows  in  air  conditioned  buildings  exclude  outside  noise,  which  makes  all  inside  noise 
from  mechanical  equipment  seem  louder. 

3  •  Name  the  acoustical  problems  which  should  be  solved  in  connection  with 
the  installation  of  heating  or  air  conditioning  equipment. 

Selection  of  quietly  operating  equipment;  adequate  insulation  of  walls  surrounding  the 
equipment  room;  mounting  of  all  vibrating  equipment  on  flexible  supports  which  will 
eliminate  solid-borne  vibrations;  design  of  suitable  sound  filters  to  reduce  the  trans- 
mission of  noise  through  ventilating  ducts;  the  use  of  suitably  low  air  speeds  and  stream- 
lining, where  necessary,  to  prevent  eddy  noises. 

4  •  Are  good  heat  insulators  also  good  sound  insulators? 

As  a  rule,  no.  Blankets  and  felted  materials  offer  considerable  insulation  for  sounds  of 
high  frequency,  but  very  little  for  sounds  of  low  frequency. 

5  •  What  is  the  principal  consideration  in  the  selection  of  elastic  supports  for 
the  insulation  of  machinery  vibration? 

The  support  should  be  so  compliant  that  the  natural  frequency  of  the  mass  of  the  machinery 
on  its  elastic  support  will  be  low  in  comparison  with  the  vibrational  frequencies  which  are 
to  be  insulated. 

6  •  What  means  should  he  utilized  for  preventing  air-borne  noise  from  the 
ventilating  equipment  from  being  transmitted  through  the  walls,  ceiling,  or 
floor  of  the  equipment  room? 

Treat  the  interior  walls  and  ceiling  of  the  equipment  room  with  absorptive  material ;  see 
that  all  doors  and  windows  to  the  equipment  room  fit  tightly  in  their  frames;  and  use 
wallr  and  floor  and  ceiling  partitions  which  have  an  insulation  value  of  not  less  than  50  db. 

313 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

7  •  Name  effective  methods  for  reducing  the  transmission  of  sound  through 
ventilating  ducts. 

Line  the  ducts  with  sound  absorptive  material,  or  use  suitable  sound  filters  made  up  of 
long  channels  of  small  cross-sectional  area,  lined  with  sound  absorptive  material. 

8  •  What  are  the  effects  of  humidity  and  temperature  on  the  absorption  of 
sound  in  air? 

The  absorption  increases  with  a  rise  in  temperature,  and  decreases  for  relative  humidities 
above  about  20  per  cent.  A  relative  humidity  of  55  to  60  per  cent  is  advantageous 
acoustically  in  large  auditoriums. 

9  •  How  may  sound  be  measured  and  what  are  the  advantages  of  the  methods 
available? 

Three  practical  methods  are  now  available  to  the  heating  and  ventilating  engineer, 
namely: 

a.  The  noise  meter  method. 

b.  The  audiometer  and  ear  method. 

c.  The  tuning  fork  and  ear  method. 

Except  for  instrument  adjustments  and  the  use  of  the  eye  in  reading  a  meter,  the  human 
element  does  not  enter  into  measurements  made  with  the  noise  meter,  so  it  is  to  be  pre- 
ferred, if  available.  The  tuning  fork  method  is  relatively  cheap  and  simple  and  suf- 
ficiently accurate  for  most  field  work.  The  audiometer  and  ear  method  ranks  between 
these  two  in  preference. 

10  •  What  are  some  of  the  more  important  sources  of  noise  in  buildings,  for 
which  the  heating  and  ventilating  engineer  may  be  held  responsible? 

a.  Furnace  room  equipment. 

b.  Radiators  and  piping. 

c .  Uncalked  openings  in  walls  around  pipes  and  ducts, 

d.  Ventilating  fans,  if  noisy  in  operation  and  not  isolated  from  the  building  structure  by 
properly  designed  vibration  damping  foundations. 

e.  High  air  velocity  in  ducts. 

/.   Ventilation  fan  rooms  not  insulated  acoustically  from  parts  of  the  building  where 

noise  would  be  objectionable, 
g.  Ventilating  ducts  without  flexible  non-metallic  sleeves  in  them  to  break  metallic 

sound  conducting  paths. 

h.  Cross  connection  of  rooms  acoustically  through  ducts. 
i.   Ventilating  ducts  without  sound  absorbing  lining,  if  required. 
j.   Unit  heaters  and  ventilators. 
k.  Unit  air  conditioners. 

11  •  The  noise  level  in  the  fan  room,  directly  under  the  main  floor  of  a  theater 
is  70  db.    The  floor  is  constructed  as  described  in  Item  5,  Table  3.    What  is  the 
fan  noise  level  in  the  theater? 

According  to  Table  3,  the  average  coefficient  of  sound  transmission,  t,  of  such  a  floor 
construction  is  0.0000020.  The  transmission  loss  through  the  floor,  expressed  in  db,  is: 

TL 


=  10  logic 


gl°  0.0000020 
57 


The  fan  noise  level  in  the  theater  would,  therefore,  be  70  db  less  57  db,  or  13  db,  which, 
according  to  Table  1,  is  an  acceptable  level. 

Another  way  of  arriving  at  the  same  result  is  by  use  of  Formula  3r  in  which  V  is  the  in- 

314 


CHAPTER  18 — SOUND  CONTROL 


tensity  of  fan  noise  as  measured  in  the  theater,  and  /"  its  intensity  as  measured  in  the 
fan  room,  I0  being  the  reference  intensity  in  both  cases,  while  -r  is  0.000002  or  2  X  10-6. 

j-  =  10' 
P 


Noise  level  =  10  logio  20  -  13  db. 


107  X  2  X  10-6  =  20 
•to 


12  •  Measurements  made  separately  of  the  noises  from  different  sources  pre- 
vailing in  a  large,  noisy  banking  room  revealed  the  following  average  noise 
levels: 

a.  From  the  street  through  windows,  doors,  and  walls,  40  db. 

b.  From  adding  machines,  typewriters,  human  movements  and  conver- 
sation, 60  db. 

c.  From  the  ventilating  system,  50  db. 
What  was  the  total  noise  level  of  the  room? 

Calling  Js,  /b,  and  /v  the  intensities  of  the  street,  banking  room,  and  ventilation  noises, 
respectively,  and  J0  the  reference  level,  we  have: 

/o  ~7o~  !o 

The  total  intensity,  I,  will  be  7S  -}-  Ib  -f  7V 

The  intensity  level  is  10  logio  -j- 

°    =101og10^dl^±_/v) 

=  10  Iog10  (104  +  106  -f-  105) 
-  60.4  db 

Note  that  the  total  loudness  level  is  not  much  above  the  level  of  the  loudest  noise.  While 
noise  intensities  may  be  added  arithmetically,  noise  levels  expressed  in  decibels  cannot 
be  so  added. 

13  •  A  ventilating  fan  room  30  ft  by  30  ft  by  12  ft  has  brick  walls,  a  concrete 
floor,  and  a  concrete  ceiling.     How  much  will  the  noise  level  of  this  room, 
expressed  in  decibels,  be  reduced  by  applying  sound  insulating  material  (co- 
efficient of  absorption  0.6  at  512  cycles)  to  two  walls  and  the  ceiling? 

Use  Formula  2: 

PS1 
I   =  before  applying  material 

PS1 
/i  —  — f-  after  applying  material 

PS1 

JL  =         a         =  J*!_ 
It  PS1  a 

a' 
Referring  to  Table  2: 

a    =  (4  X  12  X  30  X  .031)  +  (2  X  30  X  30  X  .016)  =  73.4 

a'  =  (2  X  12  X  30  X  .031)  +  (30  X  30  X  .016)  -f  (2  X  12  X  30  X  0.6)  + 
(30  X  30  X  0.6)  =  1008.7 

/    a1    1008.7 


Ji  a  73.4 


13.7 


Noise  level  reduction  =  101og1P-=-  =  10  log™  13.7  =  11.4  db. 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

14  •  What  relation  does  the  movement  for  the  suppression  of  noise  bear  to  the 
trend  toward  air  conditioning  of  offices  and  other  places  in  cities  where  people 
work  or  congregate? 

Very  important  sources  of  disturbing  sounds  are  the  various  street  noises  that  gain 
entrance,  not  only  through  open  windows  but  to  a  certain  extent  even  through  closed 
windows.  If  windows  are  to  be  kept  closed  to  exclude  noise,  air  conditioning  is  a  practical 
necessity,  especially  in  summertime.  Summertime  air  conditioning  makes  use  of 
awnings, 'which  are  not  only  desirable  but  economical  in  that  they  keep  down  cooling 
loads.  To  obviate  condensation  and  frost  on  windows,  wintertime  ^air  conditioning  calls 
for  storm  sash  or  double  glazing  which  in  turn  reduces  the  transmission  of  street  noises 


316 


Chapter  19 

AIR  DISTRIBUTION 

Warm   Air  Systems,    Combined  Systems,   Split  Systems,   School 
Buildings,   Theaters,   Upward  System,  Dowmoard  System,   Vanes 

-HTX)  produce  proper  air  distribution  in  a  room  to  be  ventilated,  heated, 
JL  or  cooled  by  air,  the  design  and  location  of  the  air  supply  inlets  and 
exhaust  outlets  must  be  carefully  considered.  A  system  may  fail  though 
it  handles  the  proper  amount  of  air  if  such  important  design  principles 
are  ignored. 

WARM  AIR  SYSTEMS 

With  gravity  warm  air  systems,  it  has  been  the  practice  to  place  the 
supply  registers  in  or  near  the  floor  of  each  room  and  to  place  the  return 
grille  in  the  floor  of  the  first  story.  When  there  is  mechanical  air  circu- 
lation, the  supply  ducts  may  be  extended  to  the  outside  walls  and  the  air 
discharged  into  the  rooms  near  their  cold  exposures;  on  the  return  side  a 
grille  is  placed  in  or  near  the  floor  at  a  central  location,  or  individual 
return  grilles  are  provided,  usually  at  the  corner  of  the  room  opposite  the 
supply  register. 

These  arrangements  are  usually  satisfactory  for  heating  (Fig.  1)  but  not 
for  cooling  (Fig.  2).  If  cool  air  is  introduced  at  one  side  of  the  room  at  the 
floor,  and  if  the  escape  opening  for  the  heated  air  to  be  displaced  by  the 
cool  air  is  at  the  floor  at  the  other  side,  the  cool  air  will  travel  across  the 
floor  and  escape  through  the  vent  or  return  air  opening,  and  thus  not 
appreciably  affect  the  warmer  air  in  the  upper  part  of  the  room. 

The  air  supply  opening  will  serve  satisfactorily  if  located  high  on  an 
interior  wall  opposite  the  exposed  wall,  and  this  location  answers  well  also 
for  gravity  indirect  heating.  The  corresponding  return  air  arrangements, 
however,  apparently  are  not  subject  to  exact  rules,  but  must  be  adapted 
to  circumstances.  For  example,  where  the  building  is  compact,  with  a 
first  story  having  rooms  open  to  each  other,  a  single,  centrally-located 
return  at  the  floor  functions  satisfactorily  for  heating,  and  if  the  second 
story  bedrooms  are  also  compactly  arranged  no  individual  return  from 
each  will  be  necessary.  On  the  other  hand,  any  room  which  is  unusually 
exposed,  which  is  especially  remote  with  reference  to  the  other  rooms,  or 
which  is  apt  to  be  tightly  closed  most  of  the  time,  should  have  a  controlled 
return  grille  and  duct.  With  a  mechanical  warm  air  system,  this  return 
may  be  close  to  the  floor,  either  below  the  supply  grille  or  under  windows 
or  other  cold  exposures,  and  with  a  gravity  system  it  may  be  close  to  the 
floor  at  the  opposite  side  of  the  room  from  the  supply  grille. 

There  is  always  an  advantage  in  keeping  the  warm  air  ducts  concen- 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


trated  nearer  the  furnace  and  not  exposing  them  to  the  influence  of  back 
drafts  of  cold  air  by  locating  them  in  outside  walls. 

COMBINED  SYSTEMS 

For  a  combined  mechanical  heating  and  cooling  system  using  refrigera- 
tion for  cooling,  no  particular  change  in  the  ducts  usually  is  necessary.  It 
is  desirable  from  an  economic  standpoint  to  take  advantage  of  the  natural 
tendency  of  the  cooler  air  to  remain  below  the  warmer  air  overhead,  and 
anything  which  will  bring  about  such  stratification  will  effect  an  economy 
in  refrigeration. 


FIG.  1.   AIR  CIRCULATION  WHEN  HEATING 
WITH  Low  SUPPLY  AND  RETURN  OPENINGS 


FIG.  2.   AIR  CIRCULATION  WHEN  COOLING 
WITH  Low  SUPPLY  AND  RETURN  OPENINGS 


FIG.  3.   AIR  CIRCULATION  WHEN  COOLING 

WITH  HIGH  SUPPLY  OPENING  AND 

Low  RETURN  OPENINGS 


FIG.  4.  SECTION  THROUGH  AW  ELEMENTAL 
MECHANICAL  WARM  AIR  HEATING- 
COOLING  SYSTEM.    THE  ATTIC 
FAN  is  ALTERNATIVE 


If  the  return  ducts  of  a  mechanically  operated  warm  air  system  are 
adequate,  appreciable  cooling  may  be  accomplished  as  follows:  The  fan 
outlet  must  have  a  by-pass  leading  to  a  basement  window  or  to  a  chimney 
provided  for  the  purpose  and  the  return  duct  must  have  an  alternative 
shaft  opening  into  the  highest  part  of  the  house.  At  night,  in  summer,  the 
fan  may  be  operated  to  exhaust  the  hot  air  from  the  top  of  the  house  by 
the  return  air  duct  just  described  and  the  fan  will  blow  this  heated  air  out 
of  doors  through  the  window,  or  preferably,  of  course,  through  the 
chimney.  The  cooler  night  air  must  then  enter  the  house  through  the 

318 


CHAPTER  19 — AIR  DISTRIBUTION 


windows,  and  by  its  motion  and  temperature  will  extract  the  heat  from 
the  walls  and  furniture.  The  cost  of  power  for  such  cooling  should  be 
carefully  checked  against  operating  with  a  much  smaller  volume  of  air 
mechanically  cooled. 

Fig.  3  shows  the  air  circulation  when  cooling  with  a  high  supply  opening 
and  a  low  return  opening.  The  air  circulation,  when  heating,  will  be 
substantially  the  same  as  when  cooling.  Fig.  4  shows  a  section  through 
an  elemental  mechanical  warm  air  heating-cooling  system.  The  attic  plan 
is  alternative.  Summer  night  cooling  may,  of  course,  be  accomplished 
by  placing  an  exhaust  fan  in  the  attic. 

SPLIT  SYSTEMS 

Many  buildings  which  are  heated  by  radiators  or  convectors  and  which 
have  rooms  requiring  ventilation  or  cooling  have  air  supply  and  exhaust 
systems  independent  of  the  radiators  or  convectors.  Such  installations 
are  termed  split  systems.  When  the  air  enters  a  room  through  conventional 
side  wall  inlets  an  occupant  may  feel  comfortable  if  the  air  is  about  the 
temperature  of  the  room,  but  the  introduction  of  too  cool  air  may  cause  a 
feeling  of  draft.  To  correct  this  draft  condition,  glass  chutes  and  elabor- 
ate diff users  are  sometimes  provided.  The  arrangement  shown  in  Fig.  5 
for  supplying  cool  air  to  a  room  provides  satisfactory  air  circulation  in 
spaces  up  to  400  sq  ft  in  area  with  ceilings  as  low  as  8  ft.  There  is  no 
maximum  ceiling  limitation  as  to  height. 

When  the  room  in  question  is  provided  with  a  unit  ventilator  which 
obtains  its  air  supply  directly  through  the  wall  from  out  of  doors,  the 
distribution  with  a  high  velocity  air  jet  passing  in  an  upward  direction 
is  quite  satisfactory. 

The  use  of  unit  air  conditioners  for  summer  cooling  introduces  no  new 
features  or  difficulties  which  have  not  already  been  encountered  in  winter 
heating.  Conditioners  must  be  provided  with  positive  control  by  means  of 
valves  or  dampers,  or  both,  which  will  prohibit  any  sudden  and  wide  tem- 
perature variation,  and  keep  the  entering  air  not  more  than  approxi- 
mately 7  deg  cooler  than  the  air  already  in  the  space.  This  temperature 
margin  is  dependent  on  various  factors  including  the  ceiling  height  of  the 
room  and  the  velocity  of  the  air  at  the  discharge  grille. 

SCHOOL  BUILDINGS 

The  air  distribution  conditions  in  school  building  classrooms  are  not 
unlike  those  illustrated  in  Fig.  1  for  mechanical  warm  air  systems  and 
those  in  Fig.  6  for  unit  ventilator  equipped  plants.  School  rooms  which 
have  center-ceiling  inlets  along  the  lines  of  Fig.  5  have  given  excellent 
results.  It  is  important  that  the  temperature  of  the  entering  air,  whether 
this  air  be  supplied  by  a  local  unit  ventilator  or  by  a  distant  central  fan, 
be  controlled  so  that  the  air  cannot  enter  the  room  from  a  side-wall  inlet 
or  from  a  unit  ventilator  at  a  temperature  more  than  a  very  few  degrees 
cooler  than  that  of  the  air  already  near  the  ceiling  of  the  room. 

Fig.  7  shows  a  section  through  a  room  equipped  with  a  unit  air  con- 
ditioner or  unit  cooler.  This  is  typical  of  the  condition  in  effect  when  any 
recirculating  room-cooling  unit  is  installed. 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

Most  unit  ventilators  employ  a  unique  method  of  air  distribution.  Its 
principal  feature  is  that  the  air  is  discharged  at  a  high  velocity  toward  the 
ceiling,  with  the  jet  inclined  slightly  toward  the  room  in  order  to  dis- 
tribute the  air  over  the  ceiling.  In  designing  a  unit  ventilator  installation 
great  pains  should  be  taken  that  nothing  will  interfere  with  the  operation 
of  this  jet.  For  this  reason  unit  ventilators  should  never  be  installed 
where  there  is  a  beam  on  the  ceiling  running  at  right  angles  to  the  direction 
of  the  air  jet.  If  ceiling  beams  cannot  be  avoided,  the  unit  ventilator 
should  be  placed  to  discharge  parallel  to  the  beams. 


wi 


A 


?  Burred  Ceding 


T 


FIG.  5.     SECTION  THROUGH   A 
RADIATOR-HEATED  ROOM 


FIG.  6.  SECTION  THROUGH  A  UNIT  VENTI- 
LATOR-EQUIPPED ROOM  WHEN  HEATING 


FIG.  7.    SECTION  THROUGH  A  UNIT  CON- 
DITIONER EQUIPPED  ROOM  WHEN  COOLING 


FIG.  8.     PLAN  OF  A  CLASSROOM  IN  A 
SCHOOL  VENTILATED  BY  A  CENTRAL  FAN 


In  Fig.  8  the  cloakroom  ceiling  is  furred  down  so  as  to  conceal  the  metal 
air  supply  duct,  which  is  close  to  the  ceiling.  The  air  for  ventilation 
usually  is  controlled  by  a  duct  thermostat  near  the  fan,  at  a  temperature 
slightly  higher  than  the  temperature  required  in  the  room,  to  allow  for 
heat  losses  in  the  duct  system. 

THEATERS 

Theaters  are  usually  ventilated  or  cooled  by  introducing  precon- 
ditioned air.  No  ventilating  system  for  a  theater  should  be  given  con- 
sideration without  definite  provision  for  cooling.  Theater  cooling 
generally  is  far  more  important  than  theater  heating.  There  are  two 
widely  different  methods  of  theater  air  distribution,  the  downward  and  the 
upward. 

320 


CHAPTER  19 — AIR  DISTRIBUTION 


Downward  System 

Theaters  usually  are  equipped  with  downward  air  distribution  with 
horizontal  diffusion  of  the  entering  cool  air  so  as  to  combine  it,  both  as  to 
temperature  and  dilution,  with  the  heated  air  which  inevitably  must  rise 
from  the  bodies  of  the  patrons.  The  waste  or  the  recirculated  air  is  with- 
drawn from  the  room  at  the  floor.  If  the  theater  is  large,  and  if  the 
exhaust  openings  are  placed  in  the  side  walls  at  the  floor,  drafts  may  be 
felt  by  the  people  who  sit  near  the  openings.  There  is  no  objection,  how- 
ever, except  that  of  cost,  to  the  use  of  small  exhaust  openings  under  each 
seat.  These  may  be  cleanable  floor  grilles  or  may  have  mushroom  covers. 

In  a  downward  system,  if  the  entering  cool  air  is  not  deflected  hori- 
zontally, it  will  fall  through  the  surrounding  much  hotter  air,  and  will 


Supply      Ducts 

— ' 


Stage 


FIG.  9.    SECTION  THROUGH  A  THEATER    FIG.  10.    THEATER  WITH  UPWARD  SYSTEM 
WITH  DOWNWARD  VENTILATION  OF  VENTILATION 


reach  high  velocities  by  the  time  it  strikes  the  heads  of  the  occupants. 
Air  at  a  temperature  10  deg  below  that  of  the  surrounding  air  is  decidedly 
objectionable  when  forced  over  one's  head  at  a  velocity  of  nearly  400  fpm. 
Fig.  9  shows  a  section  through  a  theater  with  downward  ventilation. 
The  deflectors  cause  the  entering  cool  air  to  be  spread  horizontally  so  that 
it  will  mix  with  the  hotter  air.  The  final  escape  is  through  well-distributed 
openings  in  the  floor.  There  have  been  cases  in  which  the  downward 
system  of  air  distribution  such  as  that  illustrated  in  Fig.  9  gave  trouble 
due  to  overheating  at  the  rear,  both  above  and  below  the  balcony, 
especially  when  not  provided  with  refrigeration  for  cooling,  and  when  not 
adequately  controlled.  It  is  especially  necessary  that  adequate  removal 
of  the  heated  air  be  provided  at  these  low-ceiling  points  and  it  is  probable 
that  auxiliary  exhaust  at  or  through  the  ceiling  after  the  manner  of  the 
arrangements  shown  in  Fig.  5  would  be  helpful. 

Upward  System 

If  no  inlet  openings  are  possible  in  the  ceiling,  the  upward  system  may 
be  the  less  objectionable  alternative.  Fig.  10  shows  a  section  through  a 
theater  with  the  upward  system  of  air  distribution.  The  occupants  often 
suffer  from  drafts  due  to  the  cool  air  which  comes  from  the  unoccupied 
zones. 

When  the  entire  seating  area  is  occupied,  the  upward  system  gives 
little  trouble  when  cooling,  and  since  very  little  heating  is  required  under 
such  conditions,  practically  no  difficulty  is  encountered.  The  maximum 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

volume  of  air  to  be  introduced  with  the  upward  system  is  about  25  cfm 
of  air  per  person  at  a  low  velocity,  say  at  150  fpm  (linear),  and  at  a  tem- 
perature not  more  than  6  deg  below  the  room  temperature.  For  partial 
occupancy,  higher  entering  air  temperatures  can  be  used  with  corre- 
spondingly less  danger  from  drafts. 

VANES 

In  order  to  cause  the  supply  air  to  a  room  to  take  a  fixed  or  desired 
direction  when  leaving  the  inlet  opening  of  a  flue,  stationary  vanes  may 
be  provided  at  both  the  back  of  the  grille  and  at  the  grille  to  direct  the  air 
flow.  Fig.  11  shows  a  section  through  a  room  inlet  opening  at  the  top  of  a 
rising  flue  and  indicates  the  air  conditions  when  no  vanes  are  used. 
Fig.  12  shows  a  section  through  the  same  room  inlet  opening  when  vanes 
are  advantageously  placed  to  direct  the  flow  of  air. 


FIG.    11.      Am    CONDITIONS    AT    INLET 

OPENING  AT  THE  TOP  OF  A  RISING  FLUE 

WHEN  NO  VANES  ARE  USED 


FIG.    12.      AIR    CONDITIONS    AT    INLET 

OPENING  AT  THE  TOP  OF  A  RISING  FLUE 

WHEN  DIRECTIONAL  VANES  ARE  USED 


In  many  theater  and  commercial  installations  the  ejector-like  action 
of  high-velocity  air  emerging  from  a  duct  is  taken  advantage  of,  and 
scientifically  proportioned  nozzles  are  installed  to  cause  definite  recircu- 
lation  of  the  room  air. 


PROBLEMS  IX  PRACTICE 

1  •  Is  the  conventional  warm  air  system,  employing  floor  or  baseboard  supply 
registers,  suitable  for  heating  and  cooling? 

Floor  or  baseboard  supply  registers  are  suitable  for  heating  service  because  the  natural 
tendency  of  warm  air  is  to  rise.  They  are  not  suitable  for  cooling  because  the  natural 
tendency  of  cool  air  is  to  stay  near  the  floor  and  gradually  work  its  way  to  the  return 
registers,  thus  not  cooling  the  air  in  the  upper  part  of  the  room.  See  Figs.  1  and  2. 

2  •  What  type  of  air  distribution  system  is  suitable  for  heating  and  cooling  a 
home? 

In  order  to  provide  satisfactory  cooling  without  drafts  it  is  necessary  to  discharge  the 
air  at  relatively  high  velocity  toward'  the  ceiling  from  a  high  point,  as  shown  in  Fig.  3. 
If  the  register  is  properly  designed  and  the  air  capacity  is  limited  to  approximately 
400  cfm,  the  cool  air  will  mix  with  the  air  in  the  room  before  it  drops  to  the  occupied  zone. 
However,  care  must  be  taken  that  discharged  air  does  not  impinge  on  beams  which  would 
cause  the  cool  air  to  be  deflected  downward.  This  arrangenient  is  also  satisfactory  for 
heating. 

322 


CHAPTER  19 — AIR  DISTRIBUTION 


3  •  Why  is  the  conventional  low  velocity  side  wall  inlet  unsatisfactory  for 
cooling  purposes? 

With  the  conventional  side  wall  inlet  using  velocities  of  300  to  400  fpm  the  discharged 
air  quickly  loses  its  velocity  and  drops,  causing  drafts  in  the  occupied  zone. 

4  •  \£  hat  method  of  side  wall  introduction  is  satisfactory  for  cooling  purposes 
with  a  12-ft  ceiling  height? 

The  method  shown  in  Fig.  3  can  satisfactorily  circulate  air  as  much  as  10  to  15  F  below 
room  temperature,  provided  (1}  each  jet  is  limited  to  400  cfm,  <2)  the  outlet  velocity  is 
high,  (3)  the  air  is  directed  toward  the  ceiling,  and  (4)  there  are  no  beams  on  the  ceiling. 
In  order  to  employ  this  method  in  a  classroom  it  is  usually  necessary  to  have  at  least 
three  inlets,  but  even  with  three  inlets  the  cooling  capacity  is  limited  to  that  obtained 
by  circulating  air  at  10  to  15  F  below  room  temperature. 

5  •  Should  unit  ventilators  he  considered  as  heating  units  or  as  cooling  units? 

Experience  has  shown  that  approximately  75  per  cent  of  the  time  a  classroom  is  occupied 
the  problem  is  one  of  cooling  rather  than  one  of  heating.  For  this  reason  unit  ventilators 
should  be  considered  as  cooling  units. 

6  •  What  method  of  air  distribution  is  usually  employed  with  unit  ventilators? 

Most  unit  ventilators  employ  a  unique  method  of  air  distribution  in  which  the  air  is 
discharged  at  a  high  velocity  toward  the  ceiling.  The  air  stream  is  usually  inclined 
toward  the  room. 

7  •  How  should  a  unit  ventilator  he  located  in  a.  room  that  has  ceiling  beams? 

When  there  are  ceiling  beams  the  unit  ventilator  should  be  so  located  that  the  beams  will 
be  parallel  with  the  direction  of  the  air  discharge  in  order  that  the  beams  will  not  deflect 
the  air  downward. 

8  •  Wliat  is  the  minimum  temperature  at  which  unit  ventilators  can  distribute 
air  in  a  classroom  without  causing  drafts? 

Generally  speaking,  the  lowest  minimum  discharge  temperature  at  which  objectionable 
drafts  will  not  be  created  is  60  F.  Some  designers  suggest  that  the  discharge  temperature 
can  drop  as  low  as  35  F  below  the  room  temperature  without  causing  drafts  when 
units  are  properly  installed. 

9  •  What  is  the  usual  method  of  ventilating  school  auditoriums  and  gym- 
nasiums when  unit  ventilators  are  used  in  the  classrooms? 

If  unit  ventilators  are  used  in  classrooms  the  usual  method  of  ventilating  the  auditorium 
or  gymnasium  is  to  use  one  or  more  large  units  located  above  and  on  either  side  of  the 
stage. 

10  •  What  is  the  maximum  amount  of  air  which  should  he  discharged,  from  one 
point  in  a  school  auditorium  or  gymnasium? 

The  maximum  amount  of  air  which  should  be  discharged  from  one  point  is  5000  cfm. 
This  limitation  applies  whether  the  air  is  supplied  by  units  or  by  a  central  fan  from  a 
distant  point. 

11  •  Are  vents  required  in  school  classrooms,  auditoriums,  and  gymnasiums? 

With  both  the  unit  and  the  central  fan  systems,  vents  are  usually  installed  as  a  certain 
and  positive  means  of  disposing  of  the  vitiated  and  odoriferous  air  and  also,  with  the 
central  fan  system,  for  the  further  purpose  of  effecting  a  means  of  partial  recirculation. 
Natural  outward  air  leakage  may  take  the  place  of  vents,  if  and  when  it  proves  sufficient, 
but  it  is  usually  uncertain,  insufficient,  and  uneconomical.  Vents  are  required  by  law 
in  some  communities.  If  they  are  installed,  they  should  be  provided  with  dampers  in 
order  that  they  may  be  throttled  when  required  and  closed  at  night  and  during  holidays. 

12  •  What  type  of  system  is  generally  used  in  large  continuously  operated, 
theaters? 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

Most  large  continuously  operated  theaters  are  provided  with  complete  downward 
systems  of  air  distribution  similar  to  the  one  shown  in  Fig.  9.  With  this  system  a  large 
number  of  inlet  openings  is  provided,  each  of  which  discharges  air  in  a  thin  horizontal 
stream  at  high  velocity  in  order  that  the  cool  air  will  be  mixed  with  the  air  in  the  theater 
before  it  reaches  the  patrons. 

13  •  What  system  of  air  distribution  is  frequently  used  in  smaller  theaters? 

The  system  used,  particularly  where  artificial  cooling  is  had,  brings  air  in  at  high  velocity 
through  a  large  number  of  small  horizontal  nozzles  located  in  the  rear  of  the  auditorium 
near  the  ceiling.  This  high  velocity  air  mixes  with  a  much  larger  quantity  of  air  and 
causes  circulation  within  the  theater  before  it  comes  into  contact  with  the  occupants. 
With  this  method  care  must  be  exercised  not  to  discharge  the  air  against  ceiling  beams  or 
projections  which  may  give  a  downward  direction  to  the  cool  air  before  it  is  thoroughly 
diluted. 


324 


Chapter  20 

,  AIR  DUCT  DESIGN 

Pressure  Losses.,  Friction  Losses,  Friction  Loss  Chart,  Proportioning 
the  Losses,  Sizes  of  Ducts,  General  Rules,  Procedure  for  Duct 
Design,  Air  Velocities,  Proportioning  the  Size  for  Friction,  Main 
Trunk  Ducts  with  Branches  for  Public  Buildings,  Equal  Friction 
Method,  Details  of  Duct  Construction 

THE  flow  of  air  due  to  large  pressure  differences  is  most  accurately 
stated  by  thermodynamic  formulae  for  air  discharge  under  condi- 
tions of  adiabatic  flow,  but  such  formulae  are  complicated,  and  the  error 
occasioned  by  the  assumption  that  the  gas  density  remains  constant 
throughout  the  flow  may  be  considered  negjigible  when  only  such  pressure 
differences  are  involved  as  occur  in  ordinary  heating  and  ventilating 
practice. 

In  the  development  of  the  formulae,  diagrams,  and  tables  for  the  flow 
of  air,  use  is  made  of  the  following  basic  equation  for  the  flow  of  fluids : 

If  Hv  be  the  velocity  head  in  feet  of  a  fluid,  and  the  velocity,  V,  be  expressed  in  feet 
per  minute,  the  fundamental  equation  is 


V  =  60      2g  H 


The  factor  g  is  the  acceleration  due  to  gravity,  or  32.16  ft  per  second  per  second. 

It  is  usual  to  express  the  head  in  inches  of  water  for  ventilating  work  and,  since  the 
heads  are  inversely  proportional  to  the  densities  of  the  fluids, 

#v     =    62.4 

/Zy  p 


12 
or 

Hv  =  5.2  -^ 

9 
therefore, 


V  =  1096.5  .J^X__  (1) 

I      p 
where 

V  —  velocity  in  feet  per  minute. 
hv  =  velocity  head  or  pressure  in  inches  of  water. 
p  =  weight  of  air  in  pounds  per  cubic  foot. 

For  standard  air  (70  F  and  29.92  in.  barometer)  p  =  0.07495  Ib  per  cubic  foot.    Sub- 
stituting this  value  in  Equation  1 : 


-5  V  ocfe-*  •  ^  V 


1096-5  ^^9*    =  4005  V    Av  (2) 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


70 


Llffl  ILADMS  fH  PtHCEMT  Of  PlPE 

FIG.  1.    CURVE  SHOWING  Loss  OF  PRESSURE  IN*  ROUND  ELBOWS 


The  drop  In  pressure  in  air  distributing  systems  is  due  to  the  dynamic 
losses  and  the  friction  losses.  The  friction  losses  are  those  due  to  the 
friction  of  the  air  against  the  sides  of  the  duct.  The  dynamic  losses  are 
those  due  to  the  change  in  the  direction  or  in  the  velocity  of  air  flow. 

Dynamic  Losses 

Dynamic  losses  occur  principally  at  the  entrance  to  the  piping,  in  the 
elbows,  and  wherever  a  change  in  velocity  occurs.  The  entrance  loss  Is 
the  difference  between  the  actual  pressure  required  to  produce  flow  and 
the  pressure  corresponding  to  the  flow  produced;  it  may  vary  from  0.1  to 


I' 


tOO  t5O  ZOO  .     2SO 

-LME  &ADW3  M  PE&CZNT  Of  PfPE  WlDTtt 

FIG.  2.  CURVE  SHOWING  Loss  OF  PRESSURE  IN  SQUARE  ELBOWS 

326 


CHAPTER  20 — AIR  DUCT  DESIGN 


0.5  times  the  velocity  head.  The  pressure  loss  in  elbows  must  also  be 
allowed  for  in  the  design.  It  is  customary  to  express  dynamic  losses  in 
terms  of  the  percentage  of  the  velocity  head;  in  other  words,  the  per- 
centage of  that  pressure  corresponding  to  the  average  velocity  in  the  duct 
which  is  expressed  in  terms  of  inches  of  water  gage.  Figs.  1  and  2  show 
the  effect  of  changing  the  radius  of  elbows  of  square  and  rectangular 
section.  These  charts  are  based  on  tests  of  pipe  elbows  of  ordinary  good 
sheet  metal  construction.  For  example,  a  five-piece  round  pipe  elbow 
having  a  centerline  radius  of  one  diameter  has  a  loss  of  about  25  per  cent 
of  the  velocity  head.  At  a  velocity  of  2000  fpm  the  corresponding  head 
is  0.25  in,  water  gage,  and  at  this  velocity  the  elbow  just  referred  to  would 
cause  a  pressure  drop  .of  0.063  in.  water  gage.  Experience  has  shown  that 
good  results  may  be  obtained  when  the  radius  to  the  center  of  the  elbow 
is  1J^  times  the  pipe  diameter.  The  pressure  drop  will  then  be  approxi- 
mately 17  per  cent  of  the  velocity  head  for  round  ducts,  and  9  per  cent 
for  square  ducts.  Very  little  advantage  is  gained  in  making  elbows  with 
a  radius  of  more  than  two  diameters. 

Friction  Losses 

Friction  losses  vary  directly  as  the  length  of  the  duct,  directly  as  the 
square  of  the  velocity,  and  inversely  as  the  diameter.  Since  length  is  a 
fixed  quantity  for  any  system,  the  factors  subject  to  modification  are  the 
area  and  the  velocity,  which  determine  the  relation  between  the  first  cost 
of  the  duct  system  and  the  cost  of  the  power  for  overcoming  friction. 

The  friction  between  the  moving  air  and  pipe  surface  causes  a  loss  of 
head  which  is  numerically  equal  to  the  pressure  required  to  maintain  a 
given  velocity,  and  is  expressed  in  the  following  modification  of  Fanning's 
formula: 

For  round  pipe  and  standard  air  (70  F  and  29.92  in.  barometer) 


For  rectangular  ducts 


where 

JtL  —  loss  of  head,  inches  of  water. 

(V  \2 
—  —  —  i    =  velocity  head,  inches  of  water. 
4IX/O  / 

V  =  velocity  of  air,  feet  per  minute. 
L  —  length  of  pipe  1 

D  =  diameter  of  pipe  \  all  in  feet. 

a,  b  —  sides  of  rectangular  duct    J 
/  =  coefficient  of  friction. 

C  =—  =  length  of  pipe  in  diameters  for  one  head  loss. 

For  all  practical  purposes  C  vaiies  only  with  the  nature  of  the  pipe 
surface:  C  =  60  for  perfectly  smooth  pipe;  =  55  for  pipe  as  used  in  planning 
mill  exhaust  systems;  =  50  for  heating  and  ventilating  ducts;  =  45  for 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


SOO  000 

600  ooo 

500000 
400  WO 
500000 


2000QO 


ISO 


100 


Friction  In  Inches  of  Waterper  100  Ft, 

FIG.  3.    FRICTION  OF  Am  IN  PIPES 
328 


CHAPTER  20 — AIR  DUCT  DESIGN 


smooth  and  40  for  rough  conduits  of  tile,  brick  or  concrete.  However, 
Fritzsche  states  (and  numerous  tests  check  very  closely)  that  /  varies 
inversely  as  the  2/7  power  of  the  pipe  diameter,  and  inversely  as  the  1/7 
power  of  the  velocity,  or  inversely  as  the  1/7  power  of  capacity,  which  is 
the  same  thing.  Thus  Formula  3  may  be  revised  as  follows,  based  upon  a 
loss  of  one  velocity  head  (at  2000  fpm)  in  a  length  equal  to  50  diameters 
of  24-in.  galvanized  swedged  pipe: 

L         (     V    \13/7 


The  preceding  formulae  are  based  on  standard  air,  and  for  other  con- 
ditions the  friction  varies  directly  as  the  air  density  and  inversely  (ap- 
proximately) as  the  absolute  temperature.  The  increase  of  friction  due 
to  increase  of  air  viscosity  with  increased  temperature  is  small  and  is 
generally  neglected. 

Friction  Loss  Chart 

Fig.  3  is  a  convenient  chart  for  determining  the  friction  loss  for  various 
air  quantities  in  ducts  of  different  sizes.  The  general  form  of  this  chart  is 
familiar,  but  it  should  be  noted  that  it  is  corrected  for  changes  in 
the  coefficient  of  friction  based  on  the  rule  that  the  coefficient  of  friction 
varies  inversely  as  the  2/7  power  of  the  diameter,  and  inversely  as  the 
1/7  power  of  the  velocity.  Fig.  3  is  based  on  a  loss  of  one  velocity  head 
(at  a  velocity  of  2000  fpm)  in  a  length  equal  to  50  diameters  of  24-in. 
round  galvanized-iron  duct  of  the  usual  construction.  Although  this 
chart  is  laid  out  for  a  value  of  C  equivalent  to  50,  it  may  be  used  for  other 
values  of  C  by  varying  the  friction  inversely  as  this  constant.  For  ex- 
ample, if  a  rougher  pipe  is  used  with  40  as  the  value  of  C,  the  friction  loss 

as  read  from  the  chart  should  be  multiplied  by  j^. 

Example  1.  Assume  that  it  is  desired  to  pass  10,000  cfm  of  air  through  75  ft  of  24-in. 
diameter  pipe.  Find  10,000  cfm  on  the  right  scale  of  Fig.  3  and  move  horizontally  left  to 
the  diagonal  line  marked  24-in.  The  other  intersecting  diagonal  shows  that  the  velocity 
in  the  pipe  is  3200  fpm.  Directly  below  the  intersection  it  is  found  that  the  friction  per 
100  ft  is  0.59  in.;  then  for  75  ft  the  friction  will  be  0.75  X  0.59  =  0.44  in.  In  a  like  man- 
ner any  two  variables  may  be  determined  by  the  intersection  of  the  lines  representing 
the  other  two  variables. 

Proportioning  the  Losses 

Other  losses  of  pressure  occur  at  the  entrance  to  the  duct,  through  the 
heating  units,  and  at  the  air  washer.  In  ordinary  practice  in  ventilation 
work  it  is  usual  to  keep  the  sum  of  the  duct  losses  M  to  3^  &n<i  the  loss 
through  the  heating  units  at  less  than  J^  of  the  static  pressure.  The 
remainder  is  then  available  for  producing  velocity.  In  the  design  of  an 
ideal  duct  system,  all  factors  should  be  taken  into  consideration  and  the 
air  velocities  proportioned  so  that  the  resistance  will  be  practically  equal 
in  all  ducts  regardless  of  length. 

DUCT  SIZES 

The  sizes  of  ducts  and  flues  for  gravity  or  mechanical  circulation  of  air 
are  usually  based  on  the  losses  due  to  friction,  and  these  losses  must  be 
kept  within  the  available  pressure  difference.  This  pressure  difference  in 

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CHAPTER  20 — AIR  DUCT  DESIGN 


mechanical  ventilation  is  that  derived  from  the  fan,  while  in  gravity 
ventilation  the  aspirating  effect  due  to  the  temperature  and  height  of  the 
column  of  heated  air  causes  the  pressure  difference. 

Genera!  Rules 

The  general  rules  to  be  followed  in  the  design  of  a  duct  system  are: 

1.  The  air  should  be  conveyed  as  directly  as  possible  at  reasonable  velocities  to  obtain 
the  results  desired  with  greatest  economy  of  power,  material  and  space. 

2.  Sharp  elbows  and  bends  should  be  avoided. 

3.  The  sides  of  all  ducts  or  flues  should  be  as  nearly  equal  as  possible.    (In  no  case 
should  the  ratio  between  long  and  short  sides  be  greater  than  10  to  1.) 

Procedure  for  Duct  Design 

The  general  procedure  for  designing  a  duct  system  is  as  follows: 

1.  Study  the  plan  of  the  building  and  draw  in  roughly  the  most  convenient  system  of 
ducts,  taking  cognizance  of  the  building  construction,  avoiding  all  obstructions  in  steel 
work  and  equipment,  and  at  the  same  time  maintaining  a  simple  design. 

2.  Arrange  the  positions  of  duct  outlets  to  insure  the  proper  distribution  of  heat. 

3.  Divide  the  building  into  zones  and  proportion  the  volume  of  air  necessary  to 
supply  the  heat  for  each  zone. 

4.  Determine  the  size  of  each  outlet,  based  on  the  volume  as  obtained  in  the  preceding 
paragraph,  for  the  proper  outlet  velocity. 

5.  Calculate  the  sizes  of  all  main  and  branch  ducts  by  either  of  the  following  two 
methods: 

a.  Velocity  Method.  Arbitrarily  fix  the  velocity  in  the  various  sections,  reducing  the 
velocity  from  the  point  of  leaving  the  fan  to  the  point  of  discharge  to  the  room.  In 
this  case  the  pressure  loss  of  each  section  of  the  duct  is  calculated  separately  and 
the  total  loss  found  by  adding  together  the  losses  of  the  various  sections. 

b.  Friction  Pressure  Loss  Method.   Proportion  the  duct  for  equal  friction  pressure 
loss  per  foot  of  length. 

6.  Calculate  the  friction  for  the  duct  offering  the  greatest  resistance  to  the  flow  of 
air,  which  resistance  represents  the  static  pressure  which  must  be  maintained  in  the  fan 
outlet  or  in  the  plenum  space  to  insure  distribution  of  air  in  the  duct  system.  The  duct 
having  the  greatest  resistance  will  usually  be  that  having  the  longest  run,  although  not 
necessarily  so. 

Air  Velocities 

The  following  velocities  of  air  are  considered  standard  for  public 
buildings: 

1.  Through  the  outside  air  intakes,  1000  fpm, 

2.  Through  connections  to  and  from  heating  unit,  1000  to  1200  fpm. 

3.  Through  the  main  discharge  duct,  from  1200  to  1600  fpm. 

4.  In  branch  ducts,  600  to  1000,  and  in  vertical  flues,  400  to  800  fpm. 

5.  In  registers  or  grilles,  200  to  400  fpm  depending  upon  the  size  and  location.    If 
diff users  of  proper  design  are  used,  25  per  cent  higher  air  velocities  are  permissible. 

These  duct  velocities  may  safely  be  increased  20  per  cent  if  first-class 
construction  is  used  to  prevent  any  breathing,  buckling,  or  vibration. 
High  velocities  "at  one  point  in  the  system  neutralize  the  effect  of  proper 
design  at  all  other  points;  hence  the  importance  of  splitters  in  elbows  and 
similar  precautions.  For  industrial  buildings  noise  is  seldom  considered, 
and  main  duct  velocities  as  high  as  2800  or  3000  fpm  may  be  used  where 
conditions  will  permit.  For  department  stores  and  similar  buildings, 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


e-4 
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§      £         § 


332 


CHAPTER  20 — AIR  DUCT  DESIGN 


maximum  velocities  with  good  construction  and  design  may  be  as  high 
as  2000  or  2200  fpm  in  main  ducts,  with  suitable  reduction  in  branches 
and  outlets.  With  these  velocities  first-class  duct  construction  is  essential. 

Proportioning  the  Size  for  Friction 

By  means  of  Figs.  4  and  5  the  diameter  of  branch  pipes  necessary  to 
carry  a  given  percentage  of  the  total  air  in  the  main  pipe  and  to  maintain 
equal  friction  per  foot  of  the  length  through  the  entire  system  may  be 
determined.  These  charts,  as  well  as  Fig.  3,  are  based  on  the  assumption 
that  the  coefficient  of  friction  varies  inversely  as  the  1/7  power  of  the 
capacity. 

Example  2.  Suppose  a  60-in.  main  pipe  is  to  be  used,  and  it  is  desired  to  know  the 
size  of  branch  pipe  required  to  carry  50  per  cent  of  the  total  air  in  the  main.  Find  50 
per  cent  at  the  left  of  the  chart,  move  right  to  the  60-in.  diagonal  line  and  note  directly 
above  at  the  top  of  the  chart  that  the  branch  pipe  will  be  46.5  in.  in  diameter. 

Where  rectangular  ducts  are  used  it  is  frequently  desirable  to  know  the 
equivalent  diameter  of  round  pipe  to  carry  the  same  capacity  and  have 
the  same  friction  per  foot  of  length.  Table  1  gives  directly  the  circular 
equivalent  of  rectangular  ducts  for  equal  friction  and  capacity.  To 
obtain  the  size  of  rectangular  ducts  for  different  capacities,  but  of  the 
same  friction  per  foot  of  length,  first  obtain  the  equivalent  round  pipe  for 
equal  friction.  Thus,  if  a  branch  of  sufficient  size  to  carry  30  per  cent  of 
a  12  x  36-in.  pipe  is  desired,  it  is  found  from  Table  1  that  the  main  is 
equivalent  to  a  22.2-in.  diameter  round  pipe.  From  Fig.  5,  30  per  cent  of 
this  is  a  pipe  14.3  in.  in  diameter,  and  referring  again  to  Table  1,  the 
rectangular  equivalent  branch  is  a  12  x  14-in.,  10  x  17J^-in.,  or  any  other 
desirable  combination. 

Multiplying  or  dividing  the  length  of  each  side  of  a  pipe  by  a  constant 
is  the  same  as  multiplying  or  dividing  the  equivalent  round  size  by  the 
same  constant.  Thus,  if  the  circular  equivalent  of  an  80  x  24-in.  duct  is 
required,  it  will  be  just  twice  that  of  a  40  x  12-in.  duct,  or  2  X  23.3  = 
46.6  in. 

DUCTS  FOR  PUBLIC  BUILDINGS 

A  main  duct  with  branches  is  generally  used  to  convey  tempered  air 
for  ventilation  purposes  only.  In  place  of  individual  ducts,  a  compara- 
tively large  main  duct  supplies  air  by  branches  to  the  room  or  rooms.  The 
velocities  vary  according  to  the  nature  of  the  installation  and  the  degree  of 
quietness  required.  At  the  start  of  the  run  a  velocity  as  high  as  2000  fpm 
may  be  used,  but  this  is  considered  the  maximum  for  public  building 
work,  and  is  reduced  to  from  400  to  800  fpm  in  the  risers.  This  duct  system 
may  be  designed  so  that  the  loss  of  pressure  in  the  branches  is  equalized  in 
a  manner  similar  to  that  previously  described. 

Equal  Friction  Method 

Example  S.  Fig.  6  shows  a  typical  layout  of  an  air  distribution  'system  which  is 
applicable  for  ventilation  of  hotel  dining  rooms  and  offices. 

The  volume  of  air  in  cubic  feet  per  minute  for  the  room  is  determined  on  the  basis  of 
the  number  of  air  changes  per  hour  required.  In  the  example  shown,  the  room  ventilated 
is  a  hotel  diningf  room  135  ft  x  85  ft  x  15  ft.  A  7J4-minute  air  change  (8  air  changes  per 
hour)  is  assumed  for  proper  ventilation,  giving  22,935  cfm  as  the  air  required. 

22  935 
The  clear  area  of  the  fresh  air  inlet  is  based  on  a  velocity  of  1000  fpm  or     ^       = 

333 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


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334 


CHAPTER  20 — AIR  DUCT  DESIGN 


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AMERICAN  SOCIETY  of  HEATING  dnd  VENTILATING  ENGINEERS  GUIDE,  1935 


336 


CHAPTER  20 — AIR  DUCT  DESIGN 


22.94  sq  ft.  If  the  air  washer  is  provided  with  automatic  humidity  control,  the  tempering 
coil  should  raise  the  temperature  of  the  entering  air  to  32  F.  The  washer  with  its  auto- 
matic control  will  then  raise  the  temperature  from  32  F  to  42  F.  If  the  washer  is  not 
provided  with  automatic  humidity  control,  the  tempering  coil  must  raise  the  temperature 
of  the  entering  air  to  at  least  55  F  to  allow  for  some  temperature  drop  in  the  washer  due 
to  evaporation.  The  reheating  coil  is  selected  to  raise  the  temperature  of  the  air  from 
that  leaving  the  air  washer  to  70  F.  The  air  washer  should  have  a  maximum  velocity  of 
500  fpm  through  the  clear  area,  which,  in  this  case,  is  46  sq  ft.  For  more  detailed  infor- 
mation on  tempering  coil  and  air  washer  control,  see  Chapters  23  and  14. 

Since  the  plan  shows  a  moderately  short  run  of  main  duct  with  no  risers  near  the  fan 
outlet,  a  fan  should  be  selected  which  will  have  the  required  capacity  of  22,935  cfm  with 
a  maximum  velocity  through  the  fan  outlet  of  1400  fpm.  The  outlet  area,  therefore, 
should  be  16J^  sq  ft. 

TABLE  2.    PIPE  SIZES  FOR  EXAMPLE  3a 


VOLUME 

OF  AlH 

(CFM) 

PEE  CENT 
o?  TOTAL 
|               VOLUME 

DIAMETER  or 
PIPH 

(               (INCHES) 

EQUIVALENT  SIZE  OF 
RECTANGULAR  DUCT 
(INCHES) 

22,935 

1            100.0 

J              56                 !                   60x44 

12,510 

;             54.6 

45                 ;                  58  x  30 

10,425 

45.4 

«              42 

50x30 

8,340 

'              36.3 

!             39 

42x30 

6,255 

!              27.2 

35 

42x24 

4,170 

i              18.2 

291A             ',                   30x24 

2,085 

9.1 

23 

30x15 

I 

[ 

a  Velocity  through  diffusers  (not  shown)  to  be  approximately  300  fpm. 

The  main  pipe  size  should  be  selected  to  give  a  velocity  equal  to  or  less  than  the 
velocity  at  the  fan  outlet.  Choosing  a  56-in.  pipe  with  a  cross-sectional  area  of  17.1  sq  ft, 
the  velocity  in  the  main  pipe  will  be  1340  fpm.  Using  the  friction  pressure  loss  method 
this  56-in.  main  pipe  will  be  taken  as  the  basis  of  calculation. 

Fig.  6  shows  the  amount  of  air  to  be  handled  by  each  section  of  pipe.  Expressing  the 
volume  handled  by  each  section  as  a  percentage  of  the  total  volume  and  using  the  charts, 
Figs.  4  and  5,  the  pipe  sizes  are  as  shown  in  Table  2. 

The  pressure  at  the  outlets  nearest  the  fan  will  be  greater  than  at  the  pipes  farther 
along  the  run  so  that  the  former  will  tend  to  deliver  more  than  the  calculated  amount  of 
air.  To  remedy  this  condition,  volume  regulating  dampers  should  be  located  at  the  base 
of  each  riser  and  adjusted  for  proper  distribution.  At  points  where  branches  leave  the 
main  it  may  be  advisable,  depending  upon  the  nature  of  the  installation,  to  install 
adjustable  splitters  similar  to  that  shown  in  Fig.  6  where  the  main  duct  divides  into  the 
58  in.  X  30  in.  and  50  in.  X  30  in.  branches. 

The  rectangular  equivalents  are  selected  from  Table  1  ;  the  width  to  depth  proportion 
will  be  determined  by  construction  requirements  and  ease  of  fabrication.  The  calcu- 
lation of  the  friction  is  as  follows: 

The  longest  run  from  the  fan  outlet  to  diffuser  is  150  ft  0  in.;  150  ft  of  56-in.  pipe  is 

.     ,     .  .     150  X  12  ooo,ra 

equivalent  to  -  rr  —  -  ___________________________________  —  ..................................................  ~6£»&  dia. 

*K) 

Two  45-in.,  90-deg  elbows  (2  X  g|  X  10)  ____________  .  .........................  -------  .  ................  16.1  dia. 

OQ 

Two  23-in.,  90-deg  elbows  (2  X  gg  X  10)—  ...............  ..._  .......................................    8.2  dia. 

23 
Two  23-in.,  90-deg  elbows  in  riser  (2  X  ^  X  30)  .............................................  —.  24.7  dia. 

(Two  bad  elbows  in  riser,  each  equivalent  to  30  diameters  of  duct). 


Total  diameter  of  56-in.  pipe  _______________________  ..................................................  81.2 

337 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

/1340\2 
The  velocity  head  corresponding  to  a  velocity  of  1340  fpm  is  (  TQ/VF  )     =  0.112  in. 

81  2 

Taking  50  diameters  as  one  head  loss,  then       '     X  0.112  =  0.182  in.  static  loss  in  duct. 

ou 

Where  the  connection  pieces  are  made  with  long  easy  slopes  and  the  general  work- 
manship is  good,  a  regain  in  static  pressure  may  be  deducted  from  the  foregoing  pressure 
loss.  This  can  be  taken  as  approximately  two-third?  the  difference  in  velocity  pressures 
at  the  fan  outlet  and  the  last  run  of  pipe.  The  velocity  in  the  riser  is  667  fpm  with  a 
corresponding  velocity  pressure  of  0,033  in.  The  fan  outlet  velocity  is  1400  fpm  with 
a  corresponding  velocity  pressure  of  0.122  in.  The  regain  equals  %  (0.122  —  0.033) 
=  0.059  in. 


The  net  static  pressure  loss  in  the  duct  only  is  then : 
0.182  in.  -  0.059  in 


..0.123  in. 


Other  friction  losses  are  as  follows: 

(1)  Fresh  air  intake  1000-fpm  velocity  (11A  heads  X  0.0625) 0.094  in. 

(2)  Tempering  coil  loss  (from  manufacturer's  tables) 0.100  in. 

(3)  Air  washer  loss  (from  manufacturer's  tables)... 0.250  in. 

(4)  Reheating  coil  loss  (from  manufacturer's  tables)... 0.100  in. 

(5)  Allowance  for  regulating  dampers  and  diffusers 0.100  in. 


Static  pressure  loss  of  system 0.767  in. 

The  fan  should  be  selected  from  the  manufacturer's  ratings  which,  according  to  the 
Standard  Test  Code  for  Disc  and  Propeller  Fans,  Centrifugal  Fans  and  Blowers1,  will 
deliver  22,935  cfm  at  a  static  pressure  of  0.767  in.  and  which  has  an  outlet  area  of  16H 
sqft. 

The  method  of  design  used  in  Example  3  is  the  equal  friction  method 
described  under  the  heading  Procedure  for  Duct  Design.  This  involves 
the  arbitrary  reduction  of  velocity  from  the  fan  outlet  to  the  point  of 
discharge  to  the  room,  and  the  friction  is  calculated  by  adding  the  pressure 
losses  of  each  section  of  duct.  This  method  requires  dampering  in  the 
risers. 

Example  4-  Fig.  7  shows  an  exhaust  system  layout  for  exhausting  from  buildings  of 
the  same  type  as  in  Example  3,  Assume  the  air  requirements  based  on  the  number  of 
air  changes  per  hour  to  be  16,800  cfm.  Using  a  velocity  of  1400  fpm  in  the  main  duct  at 

TABLE  3.    PIPE  SIZES  FOR  EXAMPLE  4a 


VOLUME 
or  Am 
(CFM) 

PEE  CENT 

OF  TOTAL 

VOLUME 

DIAMETER  OF 
PIPE 
(INCHES) 

EQUIVALENT  SIZE  OF 
RECTANGULAR  DUCT 
(INCHES) 

16,800 

100.0 

47 

38x48 

11,550 

68.8 

41 

30x46 

9,450 

56.2 

38 

30x40 

5,250 

31.3 

31 

24x34 

4,200 

25.0 

28.5 

24x28 

3,150 

18.8 

25.3 

16x34 

2,100 

12.5 

21.6 

16x24 

a  Velocity  through  intake  grilles  (not  shown)  to  be  approximately  400  fpm. 


*See  Chapters  17  and  41. 


338 


CHAPTER  20 — AIR  DUCT  DESIGN 


FIG.  7.    EXHAUST  SYSTEM  LAYOUT 

the  fan  inlet,  which  Is  an  average  velocity  for  this  type  of  system,  the  area  of  the  main  is 
12  sq  ft,  which  corresponds  to  a  47-in.  pipe.  Referring  to  Example  3,  and  using  the 
charts,  Figs.  4  and  5,  the  pipe  sizes  are  as  indicated  in  Table  3. 

All  risers  will  require  dampering  as  in  Example  3.    The  calculation  of  the  friction 
is  as  follows: 

The  longest  run  from  the  intake  grille  to  fan  inlet  is  100  ft. 

(TOO  ^  12\ 
-yj J 25.6  dia. 

Two  28^-in.,  90-deg  elbows  in  riser  (12<28*X80^  36  4  dia 

(Two  bad  elbows  in  riser  each  equivalent  to  30  diameters  of  duct). 

/ *?S  f\  \f   "\(Jf\ 

One  28H-in.,  90-deg  elbow  in  horizontal  run  ^     '  4? J 6.0  dia. 

Total  diameter  of  47-in.  pipe — - 68-0  dia. 

(1400\2 
|OQg  \    =  122  in. 

AS    "^  fl  1 04? 

Taking  50  diameters  as  one  head  loss,  then ^  '    - 0.166  in. 

(2)  Intake  loss  from  griU^(lK  heads  at  a  400  fpm  velocity  IK  X  0.01) 0.015  in. 

(3)  Static  pressure  required  to  produce  one  velocity  head  at  1400  fpm — 0.122  in. 

(4)  Loss  occasioned  by  step-up  of  velocity  (0.20  X  0.122) _ 0.024  in. 

(Ibis  loss  varies  from  0.05  to  0.40  velocity  bead  depending  upon  tne  nature  of  the  change. 
Far  average  systems  0.20  velocity  head  is  a  dose  approximation.)  

Static  pressure  loss  on  inlet  side, . 0-327  in. 

339 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


FIG.  8.     ISOMETRIC  VIEW  OFDucx 

SHOWING  LOCATION  OF  STIFFENING 

SEAMS  ON  TOP  AND  SIDE  PANELS 

OF  DUCT 


THESE  CROSSBREAKS 
"ARE  NEVER 
SHOWN  OM  A  PLAN 


SECTION 

r 

MEET 


ELEVATION 


REINFORCED 
CROSS  SEAMS 


SEAMS  BETWEEN  ADJACENT 
PANELS  OR  PLAIN  CROSS  SEAMS 


FIG.  10.    METHOD. OF  INSTALLING 
HEATING  UNIT 


FIG.  9.    DETAILS  OF  SEAMS 


FIG.  12.    FAN  DISCHARGE  CONNECTION 


FIG,  11.    INSTALLATION  OF  EASEMENT 
IN  DUCT  AROUND  OBSTRUCTION 


340 


CHAPTER  20 — AIR  DUCT  DESIGN 


To  this  must  be  added  the  resistance  on  the  discharge  side  of  the  fan.  A  fan  outlet 
velocity  of  approximately  1500  to  1000  fpm  may  be  used.  Assuming  the  fan  outlet  to 
be  equivalent  in  area  to  a  45-in.  pipe,  the  velocity  is  1525  fpm. 

Loss  on  discharge  (15  ft  from  fan  outlet  to  discharge): 

15  X  12        ... 

— —  =  4  diameters  of  4o-m.  pipe. 

'iO 

The  velocity  head  corresponding  to  a  velocity  of  1525  fpm  is  0.145  and  the  discharge- 
side  loss  is  — — gg =  0.012  in.  The  total  static  pressure  loss  of  the  system  is  then: 

0.012  -j-  0.327  =  0.339  in. 

The  fan  will  be  selected  to  handle  16,800  cfm  at  a  static  pressure  of  0.339  in.  and 
to  have  an  outlet  velocity  of  1525  fpm.  Outlet  area  11  sq  ft. 

Where  there  are  one  or  more  ducts  with  branches,  the  velocity  of  air  in 
the  ducts  may  be  either  chosen  arbitrarily  or  calculated  for  friction  losses. 
When  arbitrary  values  are  assigned,  a  certain  amount  of  dampering 
should  be  provided  for;  this  will  be  small  when  the  method  chosen  permits 
a  drop  in  velocity  as  the  quantity  of  air  is  reduced. 

After  the  total  air  quantity  and  the  size  of  fan  are  ascertained,  the  main 
duct  is  usually  fixed  as  being  at  least  equal  in  area  to  the  fan  outlet,  or 
perhaps  10  per  cent  greater.  From  this  main  pipe  all  others  are  propor- 
tioned. For  example,  if  the  main  duct  is  30  in.  in  diameter,  a  branch  to 
carry  10  per  cent  of  the  total  capacity  should  be  12.7  in.  in  diameter  (see 
Fig.  4)  in  order  to  have  the  same  friction  per  foot  of  length,  while  one 
carrying  one-half  the  total  capacity  of  a  30-in.  main  with  the  same  friction 
loss  per  foot  would  be  23.4  in.  in  diameter.  By  this  method  of  equalizing 
friction  it  is  unnecessary  to  consider  the  resistance  of  each  section  of  pipe 
independently,  but  only  to  know  the  distance  from  the  fan  outlet  to  the 
end  of  the  longest  run  of  pipe,  the  number  and  size  of  elbows,  and  the 
diameter  and  velocity  in  the  largest  pipe. 

Example  5.  If  the  greatest  length  of  piping  in  a  system  is  130  ft  with  a  26-in.  diameter 
main  pipe  and  one  20-in.  elbow,  the  piping  having  been  designed  for  equal  friction  per 
foot  of  length,  the  friction  would  be  the  same  as  for  130  linear  feet  of  26-in.  pipe,  or 
60  diameters.  To  this  should  be  added  the  friction  loss  in  elbows,  in  this  case  one  20-in. 
elbow,  which  has  a  loss  equivalent  to  one-fifth  of  a  velocity  head  or  ten  diameters  of 

20 
20-in.  pipe.   This  in  turn  is -^  X  10  =  7.7  diameters  of  26-in.  pipe.     The  total  equivalent 

length  of  the  system  will  then  be  60  -f-  7.7,  or  67.7  diameters.    Since  50  diameters  is 

f\7  7 

equivalent  to  one  velocity  head,  the  loss  is      '     =  1.35  times  the  velocity  head.     If 

ou 

the  velocity  is,  for  example,  2200  fpm,  corresponding  to  0.3-in.  pressure,  the  friction  loss 
of  the  system  will  be  1.35  X  0.3  =  0.405  in. 

Frequently  the  prevention  of  sound  in  a  heating  or  ventilating  system 
imposes  more  severe  restrictions  than  the  prevention  of  excessive  pressure 
drop.  This  question  is  highly  involved  and  requires  consideration  of 
many  factors.  The  air  velocities  to  be  used  will  vary  with  the  standard  of 
construction  used  in  the  ducts  themselves  as  well  as  with  the  nature  of  the 
occupancy  and  the  construction  of  the  building.  In  general,  architects 
and  engineers  who  leave  the  details  of  duct  construction  to  the  contractor 
must,  of  necessity,  design  for  lower  velocities  than  might  be  required  for 
quiet  operation  if  proper  construction  details  were  always  followed.  The 

341 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


ROSETTE 


FIG.  13.   AIR  SPLITTERS 
INSTALLED  IN  ELBOW 


VANES 


FIG.  14.    AIR  SPLITTERS  IN- 
STALLED IN  ELBOW  AT  FAN 
DISCHARGE 


FIG.  15.    AIR  SPLITTERS 

IN  BRANCH  DUCTS  AND 

ELBOWS 


contractor  may  be  expected  to  build  the  ducts  by  the  least  expensive 
methods,  and  the  engineer  must  anticipate  this.  For  further  information 
on  noise  reduction,  see  Chapter  18. 

Details  of  Duct  Construction 

If  panel  construction  is  used  with  standing  seams  or  similar  reinforce- 
ment, and  the  panels  are  cross-broken  to  give  rigidity,  there  is  less  like- 
lihood of  vibration  due  to  air  flow,  or  deflection  due  to  air  pressure. 
Elbows  made  without  splitters,  and  improperly  shaped  transformation 
sections  produce  high  local  velocities  which  are  the  cause  of  noise  in  duct 
work.  The  use  of  first-class  duct  construction  with  well-designed  trans- 
formation sections  and  splitters  in  elbows  tends  to  maintain  relatively 
uniform  velocities  with  decrease  in  turbulence  and  in  the  noise  produced. 

Figs.  8  to  15  show  acceptable  construction  details  for  rectangular 
ducts,  elbows,  transformation  pieces  or  connections,  and  air  splitters. 
Other  methods  are  also  acceptable,  such  as  the  use  of  angle  iron  stiffeners 
for  large  ducts.  Good  construction  is  essential  to  the  elimination  of  duct 
noises  and  for  the  prevention  of  a  flimsy  installation. 

Fig.  8  is  an  isometric  view  of  a  duct  showing  the  location  of  the 
stiffening  seams  on  the  top  and  side  panels.  The  cross  seams  should  not 
occur  at  the  same  place  but  should  be  staggered  as  indicated.  Heating 
units  should  be  installed  as  shown  in  Fig.  10  with  the  duct  connections 
making  an  angle  of  not  less  than  45  deg,  but  preferably  60  deg.  Fan  dis- 

TABLE  4.    SHEET  METAL  GAGES  FOR  RECTANGULAR  DUCT  CONSTRUCTION2- 


GA&B 

WIDTH  or  DUCT 

SEAM 

RTOWORCBD  SEAM 

26 

Up  to  12  in. 

24 

13  in.  to  30  in. 

1 

22 

31  in.  to  48  in. 

1 

22 

49  in.  to  60  in. 

1M 

J^  in.  x  1%  in. 

20 

61  in.  to  90  in. 

ii4 

Min.xlJiin. 

« If  panels  are  not  cross-broken  two  gages  heavier  material  should  be  used. 

342 


CHAPTER  20 — AIR  DUCT  DESIGN 


charge  connections  should  have  a  maximum  slope  of  1  in  7,  as  indicated  in 
Fig.  12.  Whenever  a  pipe  or  other  obstruction  passes  through  a  duct 
an  easement  should  be  placed  around  the  pipe  as  indicated  in  Fig.  11. 
Air  splitters  should  be  installed  in  elbows  as  shown  in  Figs.  13  and  14. 
The  recommended  gages  for  rectangular  sheet  metal  duct  construction  are 
given  in  Table  4. 

REFERENCES 

Fan  Engineering,  Buffalo  Forge  Co. 

Heat  Power  Engineering  by  Barnard,  Ellenwood,  and  Hirshfeld,  Part  III. 
Mechanical  Engineers'  Handbook  by  Lionel  S.  Marks,  McGraw-Hill  Book  Co. 
The  Flow  of  Liquids,  by  W.  H.  McAdams,  Refrigerating  Engineering,  February,  1925,  p.  279. 
A  Study  of  the  Data  on  the  Flow  of  Fluids  in  Pipes,  by  Emory  Jvemler,  A.S.M.E.  Transactions,  Hy- 
draulics Section,  August,  31,  1933,  p.  7. 


PROBLEMS  IN  PRACTICE 

1  •  Why  is  it  desirable  to  make  elbows  with  a  radius  equal  to  one  and  one-half 
times  the  pipe  diameter? 

Reference  to  Figs.  1  and  2  will  show  that  while  the  loss  of  velocity  head,  as  indicated  by 
the  curves,  shows  considerable  variation  for  elbows  between  the  range  of  50  and  150  per 
cent  radius,  the  line  is  practically  straight  after  150  per  cent,  indicating  very  little 
variation  in  loss  of  head  for  elbows  of  larger  radius. 

2  •  What  is  the  best  shape  to  use  for  duets? 

The  shapes  to  be  used  in  designing  ducts,  in  the  order  of  their  preference,  are  round, 
square,  and  rectangular. 

3  •  What  determines  which  shape  to  use? 

Structural  and  space  conditions.  Because  ducts  are  as  a  rule  part  of  the  building  or 
structure,  it  is  necessary  to  proportion  their  sizes  to  fit  the  spaces  available. 

4  •  What  is  meant  by  "arbitrarily  fix  the  velocity  in  the  various  sections?" 

When  using  the  vejocity  method  as  a  basis  for  design,  the  maximum  allowable  velocity 
is  fixed  for  the  main  supply  duct  at  the  fan,  and  this  velocity  is  gradually  decreased  as 
each  branch  or  outlet  is  taken  off  the  main  supply  duct. 

5  •  Which  system  of  duct  design  is  to  be  preferred,  the  velocity  method  or  the 
friction  pressure  loss  method? 

The  friction  pressure  loss  method  can  be  used  to  advantage  where  no  structural  or 
building  conditions  limit  the  shape  of  the  ducts.  Where  these  limiting  conditions  exist 
the  velocity  method  is  to  be  preferred. 

6  •  Are  the  grille  sizes  figured  on  the  same  basis  as  the  outlets? 

The  free  area  through  the  grilles  is  figured  the  same  as  the  outlets,  and  this  area  is 
increased  from  20  to  50  per  cent,  depending  on  the  design  of  the  grille,  to  allow  for  the 
loss  of  area  caused  by  the  construction  of  the  face  of  the  grille, 

7  •  Where  it  is  necessary  to  provide  steel  angle  braces,  how  far  apart  should 
they  he  spaced? 

Angle  braces  for  large  ducts  should  be  placed  on  3-ft  0-in.  centers. 

343 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

8  •  How  much  air  will  a  10-in.  by  24-in.  duct  handle  if  it  is  part  of  a  system 
designed  on  a  pressure  drop  of  0.1  in.  per  100  feet  of  run? 

1450  cfm  ( Table  1  and  Fig,  3j. 

9  •  How  does  a  splitter  at  a  duct  junction  influence  the  volume  of  the  air  going 
through  each  branch? 

A  splitter  facing  the  direction  of  air  flow  cuts  off  the  air  and  delivers  the  desired  amount 
to  the  branch. 

10  •  Why  does  a  wide,  shallow  duct  offer  more  resistance  to  the  now  of  ah*  than 
does  a  square  duct  of  equal  cross-sectional  area? 

The  perimeter  of  the  wide,  flat  duct  is  greater  than  that  of  the  square-section  duct,  so  the 
former  has  the  greater  frictional  area  which  increases  the  resistance  and  thus  reduces  the 
volume  at  any  given  pressure. 

11  •  What  methods  are  used  to  keep  large  ducts  from  vibrating  because  of  air 
pulsations,  and  from  sagging  because  of  their  own  weight? 

External  bracing,  such  as  standing  seams,  or  structural  shapes,  like  tees  or  angles,  should 
be  placed  across  the  top  and  bottom.  Exterior  braces  or  cross  buckling  of  metal  sheets 
in  diagonal  panels  may  be  used  for  the  sides  of  large  ducts. 

12  •  What  velocities  of  air  flow  should  be  used  in  the  trunk  ducts  of  a  venti- 
lating system  in  a  public  building? 

From  1200  to  1600  fpm. 

13  •  In  a  ventilating  system  in  a  residence,  what  is   the  recommended  air 
velocity  through  supply  registers  and  grilles? 

400  fpm. 


344 


Chapter  21 

E\T»USTRIAl,  EXHAUST  SYSTEMS 

Types,   Design  of  Systems,   Suction  and   Velocity  Requirements, 

Design  of  Hoods,  Design  of  Duct  Systems,  Collectors,  Resistance  of 

Systems,  Selection  of  Fans  and  Motors 

T7  XHAUST  and  collecting  systems  are  found  in  almost  every  industry 
F^  and  are  a  vital  adjunct  in  maintaining  safe  and  hygienic  conditions1. 
The  present  chapter  attempts  to  give  general  information  relating  to  the 
design  of  factory  exhaust  systems  in  order  that  efficient  and  economical 
control  of  dusts  and  fumes  may  be  achieved. 

TYPES  OF  SYSTEMS 

There  are  two  general  arrangements,  the  central  and  the  group  systems. 
In  the  central  system  a  single  or  double  fan  is  located  near  the  center  of 
the  shop  with  a  piping  system  radiating  to  the  various  machines  to  be 
served.  In  the  group  system,  which  is  sometimes  employed  where  the 
machines  to  be  served  are  widely  scattered,  small  individual  exhaust  fans 
are  located  at  the  center  of  the  machine  groups.  The  group  arrangement 
has  the  advantage  of  flexibility. 

Exhaust  systems  are  also  classified  by  the  means  employed  to  collect 
dust  or  other  material  handled.  The  dust  or  refuse  may  be  collected  and 
controlled  by  enclosing  hoods,  open  hoods,  inward  air  leakage,  or  by 
exhausting  the  general  air  of  the  room. 

With  some  classes  of  machinery  it  is  not  feasible  to  closely  hood  the 
machines  and  in  these  cases  open  hoods  over  or  adjacent  to  the  machines 
are  provided  to  collect  as  much  as  possible  of  the  dust  and  fumes.  This 
class  includes  such  machines  as  rubber  mills,  package  filling  machinery, 
sand  blast,  crushers,  forges,  pickling  tanks,  melting  furnaces,  and  the 
unloading  points  of  various  types  of  conveyors. 

The  open  hoods  should  be  placed  as  close  to  the  source  of  dust  or  fumes 
as  possible,  with  due  regard  to  the  movements  of  the  operator.  When  the 
hood  must  be  placed  at  some  distance  above  the  machine  it  should  be 
large  enough  to  encompass  an  area  of  considerable  extent  as  diffusion  is 
usually  quite  rapid. 

Consideration  must  also  be  given  to  the  natural  movement  of  the 
fumes.  For  those  that  are  lighter  than  air  the  hood  should  be  over  or 
above  the  machine  and  where  a  heavy  vapor  or  dust-laden  air  at  ordinary 
temperature  is  to  be  removed,  horizontal  or  floor  connections  are  required. 
If  it  is  attempted  to  remove  heavy  dust  such  as  lead  oxides  by  an  over- 
head hood  the  conditions  may  be  worse  than  if  no  exhaust  were  used  at 


Criteria  for  Industrial  Exhaust  Systems,  by  J.  J.  Bloomfield  (A.S.H.V.E.  Journal  Section,  Heating, 
Piping  and  Air  Conditioning,  July,  1934). 

345 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

all,  owing  to  the  rising  air  current  carrying  the  dust  up  through  the 
breathing  zones.  The  objective  to  keep  in  mind  in  all  cases  is  to  take 
advantage  of  the  natural  tendency  of  the  material  to  move  upward  or 
downward. 

In  another  class  of  operation  the  main  objective  is  to  prevent  the  escape 
of  dust  into  the  surrounding  atmosphere,  the  removal  of  some  dust  from 
the  machine  or  enclosure  being  merely  incidental.  The  dust-creating 
apparatus  is  enclosed  within  a  housing  which  is  made  as  tight^  as  prac- 
ticable, and  sufficient  suction  is  applied  to  the  enclosure  to  maintain  an 
inward  air  leakage,  thus  preventing  escape  of  the  dust.  While  the  exhaust 
system  is  required  to  handle  only  the  air  which  leaks  in  _ through  the 
crevices  and  openings  in  the  enclosure,  yet  in  many  installations  leakages 
are  very  high  and  great  care  is  required  to  obtain  satisfactory  results 
with  a  system  of  this  kind.  The  inward-leakage  principle  is  utilized  for 
controlling  dust  in  the  operating  of  tumbling  barrels,  grinding,  screening, 
elevating,  and  similar  processes. 

Certain  dust  and  fume  producing  operations  are  best  carried  on  by 
isolating  the  process  in  a  separate  compartment  or  room  and  then  apply- 
ing general  ventilation  to  this  space.  The  compartment  or  room  in  which 
the  work  is  performed  should  be  as  small  as  is  consistent  with  convenience 
in  handling  the  work.  The  ventilating  system  should  be  designed  so 
that  a  strong  current  of  clean  air  is  drawn  across  the  operator,  and  away 
from  him  toward  the  work,  where  the  dust  is  picked  up  and  carried 
from  the  room. 

DESIGN  OF  SYSTEMS 

The  first  step  in  the  design  of  an  exhaust  system  is  to  determine  the 
number  and  size  of  the  hoods  and  their  connections.  No  general  rules, 
however,  can  be  given  since  hood  and  duct  dimensions  are  determined  by 
the  characteristics  of  the  operations  to  which  they  are  applied.  When  a 
tentative  decision  regarding  the  set-up  has  been  made,  it  is  then  necessary 
to  obtain  the  suction  and  air  velocities  required  to  effect  control.  At  this 
point  the  designer  must  rely  upon  the  prevailing  practice  and  on  such 
physical  data  relating  to  hoods,  duct  systems  and  collectors  as  are  avail- 
able. Finally,  in  choosing  the  fan,  the  area  of  the  intake  should  be  equal 
to  or  greater  than  the  sum  of  the  areas  of  the  branch  ducts.  The  speed,  of 
course,  must  be  sufficient  to  maintain  the  estimated  suction  and  air 
velocities  in  the  system.  In  general,  the  most  important  requirements  of 
an  efficient  exhaust  and  collecting  system  are  as  follows2 : 

1.  Hoods,  ducts,  fans  and  collectors  should  be  of  adequate  size. 

2.  The  air  velocities  should  be  sufficient  to  control  and  convey  the  materials  collected. 

3.  The  hoods  and  ducts  should  not  interfere  with  the  operation  of  a  machine  or  any 
working  part. 

4.  The  system  should  do  the  required  work  with  a  minimum  power  consumption. 

5.  When  inflammable  dusts  and  fumes  are  conveyed,  the  piping  should  be  provided 
with  an  automatic  damper  in  passing  through  a  fire-wall. 

6.  Ducts  and  all  metal  parts  should  be  grounded  to  reduce  the  danger  of  dust  ex- 
plosions by  static  electricity. 

7.  The  design  of  an  exhaust  system  should  afford  easy  access  to  parts  for  inspection 
and  care. 


2For  more  detailed  requirements  see  Safe  Practice  Pamphlets  Nos,  32  and  37,  published  by  thtNaifonel 
Safety  Council,  Chicago. 

346 


CHAPTER  21 — INDUSTRIAL  EXHAUST  SYSTEMS 


SUCTION  AND  VELOCITY  REQUIREMENTS 

The  removal  of  dust  or  waste  by  means  of  an  exhaust  hood  requires  a 
movement  of  air  at  the  point  of  origin  sufficient  to  carry  It  to  a  col- 
lecting system.  The  air  velocities  necessary  to  accomplish  this  depend 
upon  the  physical  properties  of  the  material  to  be  eliminated  and  the 

TABLE  1.    SIZE  OF  CONNECTIONS  FOR  WOOD- WORKING  MACHINERY 


TYPE  OF  MACHINE 


DIAMETER  OF 
CONNECTIONS  IN 

INCHES 


Circular  saws,  12-in.  diam — ;  4 

Circular  saws,  12-24-in.  diam I  5 

Circular  saws,  24-40-in.  diam ;  6 

Band  saws,  blade  under  2  in.  wide._ 4 

Band  saws,  blade  2-3  in.  wide._ „ 5 

Band  saws,  blade  3-4  in.  wide j  6 

Band  saws,  blade  4-5  in.  wide J  7 

Band  saws,  blade  5-6  in.  wide._ '  8 

Small  mortisers j  6 

Single  end  tenoners „ j  6 

Double  end  tenoners _ !  7 

Double  end,  double  head  tenoners _ „ 10 

Planers,  matchers,  moulders,  stickers,  jointers,  etc. — 

With  knives,    6-10  in „ 5-6 

With  knives,  10-20  in 6-8 

With  knives,  20-30  in .__ „ „ 6-10 

Shapers,  light  work j  4—5 

Shapers,  heavy  work _ j  8 

Belt  sander,  belt  less  than  6  in.  wide._ 5 

Belt  sander,  belt    6-10  in.  wide „ 6 

Belt  sander,  belt  10-14  in.  wide I 7 

Drum  sander,  24  in „ 5 

Drum  sander,  30  in. _ „ 6 

Drum  sander,  36  in , 7 

Drum  sander,  48  in. 8 

Drum  sander,  over  48  in 10 

Disc  sander,  24  in.  diam. 5 

Disc  sander,  26-36  in.  diam. . 6 

Disc  sander,  36-48  in,  diam 7 

Arm  sander _ _ 4 


direction  and  speed  with  which  it  is  thrown  off.  If  the  dust  to  be  removed 
is  already  in  motion,  as  is  the  case  with  high-speed  grinding  wheels,  the 
hood  should  be  installed  in  the  path  of  the  particles  so  that  a  minimum 
air  volume  may  be  used  effectively.  It  is  always  desirable  to  design  and 
locate  a  hood  so  that  the  volume  of  air  necessary  to  produce  results  is  as 
small  as  possible. 

The  static  suction  at  the  throat  of  a  hood  is  frequently  used  in  practice 
as  a  measure  of  the  effectiveness  of  control*  This  is  of  considerable  value 
where  exhaust  systems  adapted  to  particular  operations  have  been 
standardized  by  practice.  Tables  1  and  2  present  the  duct  sizes  usually 
employed  for  standard  wood-working  machinery  and  for  grinding  and 
buffing  wheels.  Static  pressures  which  in  practice  have  been  found 
necessary  to  control  and  convey  various  materials,  are  given  in  Table  3. 
It  must  be  remembered,  however,  that  the  suction  is  merely  a  rough 

347 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


TABLE  2.    SIZE  OF  CONNECTIONS  FOR  GRINDING  AND  BUFFING  WHEELS 


MAX. 

MlN.  DlAM. 

DIAMETER  < 

3F  WHEELS 

GRINDING 
SURFACE 

OF  BRANCH 
PIPES  IN 

SQ  IN. 

INCHES 

Grinding  — 

6  in.  or  less,  not  over  1 

in.  thick. 

19 

3 

7  in.  to    9  in.  inclusive 

,  not  over  IJ' 

^  in.  thick  

43 

10  in.  to  16  in,          * 

u         u      2 

in.      a    

101 

4 

17  in.  to  19  in.          " 

a         «      3 

in.      " 

180 

4J^ 

20  in.  to  24  in.          a 

*       «     4 

in.      «    ._.. 

302 

5 

25  in.  to  30  in.          a 

«       u     5 

in.      a    _„ 

472 

6 

Buffing  — 

6  in.  or  less,  not  over  1 

in.  thick.    

19 

3V*> 

7  in.  to  12  in.  inclusive 

,  not  over  IJ/ 

^  in.  thick  

57 

4 

13  in.  to  16  in.          " 

a         «      2 

in.      "    ..„„ 

101 

4J^ 

17  in.  to  20  in. 

a         «      3 

in.      u    

189 

5 

21  in.  to  27  in.          « 

«         «      4 

in.      a    

338 

6 

27  in.  to  33  in.          tf 

«       «     5 

in.      a 

518 

7 

TABLE  3.    SUCTION  PRESSURES  REQUIRED  AT  HOODS 


STATIC  SUCTION  IN 
INCHES  OF  WATER 


Exhausting  from  grinding  and  buffing  wheels 

Exhausting  from  tumbling  barrels 

Exhausting  from  wood-working  machinery — light  duty 

Exhausting  from  wood-working  machinery — heavy  duty 

Shoe  machinery  exhaust 

Exhausting  from  rubber  manufacturing  processes 

Flint  grinding  exhaust . 

Exhausting  from  pottery  processes..... 

Lead  dust  and  fume  exhaust 

Fur  and  felt  machinery  exhaust-- 


Exhausting from  textile  machinery. 

Exhausting  from  elevating  and  crushing  machinery 

Conveying  bulky  and  heavy  materials 


2 

2 

2-4 

2-3 

2 

2  ' 

2 

2-4 

2-3 

2-3 

2 

3-5 


measure  of  the  air  volume  handled  and  consequently  of  the  air  velocity  at 
the  opening  of  the  hood.  The  elimination  of  any  dusty  condition  requires 
added  information  concerning  the  shape,  size  and  location  of  the  hood 
used  with  regard  to  the  operation  in  question. 

In  some  states  grinding,  polishing  and  buffing  wheels  are  subject  to 
regulation  by  codes.  The  static  suction  requirements,  which  range  from 
1^4  to  5  in.  water  displacement  in  a  £/-tube,  should  be  followed  although 
in  several  instances  they  may  appear  to  be  excessive.  Frequently,  in 
these  operations,  a  large  part  of  the  wheel  must  be  exposed  and  the  dust- 
laden  air  within  the  hood  is  thrown  outward  by  the  centrifugal  action  of 
the  wheel,  thus  counteracting  useful  inward  draft.  This  tendency  may 
be  diminished  by  locating  the  connecting  duct  so  as  to  create  an  air  flow 
of  not  less  than  200  fpm  about  the  lower  rim  of  the  wheel. 

Exact  determinations  of  hood  control  velocities  are  not  available,  but 

348 


CHAPTER  21 — INDUSTRIAL  EXHAUST  SYSTEMS 


It  is  safe  to  assume  that  for  most  dusty  operations  they  should  not  be  less 
than  200  fpm  at  the  point  of  origin.  For  granite  dust  generated  by 
pneumatic  devices,  Hatch  et  al3  give  velocities  from  150  to  200  fpm, 
depending  on  the  type  of  hood  used,  as  sufficient  for  safe  control.  Con- 
sidering the  character  of  the  industry,  air  velocities  of  this  order  may  be 
extended  to  similar  dusty  operations.  The  method  for  approximately 
determining  these  velocities  in  terms  of  the  velocity  at  the  hood  opening 
is  given  below. 

DESIGN  OF  HOODS 

No  set  rule  can  be  given  regarding  the  shape  of  a  hood  for  a  particular 
operation,  but  it  is  well  to  remember  that  its  essential  function  is  to  create 
an  adequate  velocity  distribution.  The  fact  that  the  zone  of  greatest 
effectiveness  does  not  extend  laterally  from  the  edges  of  the  opening  may 
frequently  be  utilized  in  estimating  the  size  of  hood  required.  Where 
complete  enclosure  of  a  dusty  operation  is  contemplated,  it  is  desirable  to 
leave  enough  free  space  to  equal  the  area  of  the  connecting  duct.  Hoods 
for  grinding,  polishing  and  buffing  should  fit  closely,  but  at  the  same  time 
should  provide  an  easy  means  for  changing  the  wheels.  It  is  advisable  to 
design  these  hoods  with  a  removable  hopper  at  the  base  to  capture  the 
heavy  dusts  and  articles  dropped  by  the  operator.  Such  provisions  are  of 
assistance  in  keeping  the  ducts  clear.  Air  volumes  used  to  control  many 
dust  discharges  may  often  be  reduced  by  effective  baffling  or  partial 
enclosure  of  an  operation.  This  procedure  is  strongly  urged  where  dusts 
are  directed  beyond  the  zone  of  influence  of  the  hood. 

Axial  Velocity  Formula  for  Hoods 

When  the  normal  flow  of  air  into  a  hood  is  unobstructed,  the  following 
formula  may  be  used  to  determine  the  air  velocity  at  any  point  along  the 
axis: 


100  -  Y         ** 

where 

Y  —  per  cent  of  velocity  at  opening. 
A  =  area  of  opening,  square  inches  (or  square  feet). 
x  =  distance  outward  from  opening,  inches  (or  feet). 

It  is  important  to  note  that  the  velocity  function  varies  in  direct 
proportion  to  the  area.  Hence,  under  certain  conditions,  a  large  opening 
may  function  more  effectively  than  a  small  one  for  the  same  volume  of 
flow.  The  formula,  of  course,  presumes  that  the  air  velocity  distribution 
across  the  hood  opening  is  uniform4. 

Example  1.  A  small  hood  64  sq  in.  in  area  handles  400  cfm.  What  will  be  the  air 
velocity  at  a  point  5  in.  outward  along  the  axis  if  the  flow  is  unobstructed? 


*Hatch,  Theodore,  Drinker,  Philip,  and  Choate,  Sarah  P.,  Control  of  the  SiEcosis  Hazard  in  the  Hard 
Rock  Industries.  I.  A  Laboratory  Study  of  the  Design  of  Dust  Control  Systems  for  Use  with  Pneumatic 
Granite  Cutting  Tools.  (Journal  of  Industrial  Hygiene,  VoL  XII,  No.  3,  March,  1930). 

^Velocity  Characteristics  of  Hoods  under  Suction,  by  J.  M.  DaHaVaHe  (A.S.H.V.E.  TRANSACTIONS 
Vol.  38,  1932). 

349 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1035 

Solution.    Substitute  in  Equation  1  and  solve  for  F,  thus 

Y  0.1  X  64 

100  -  F          5X5 

from  which  Y  —  20.4  per  cent  of  the  velocity  at  the  opening  of  the  hood. 

400  X  144 
Velocity  at  opening  =  ^ ~  900  fpm 

Hence,  the  velocity  at  the  point  in  question  is  900  X  0.204  =  184  fpm 

Air  Flow  from  Static  Readings 

The  volume  of  air  flow  into  any  hood  may  be  determined  from  the 
following  equation : 

Q  *  4005  fa  V/zT  (2) 

where 

Q  —  volume  of  air  flow,  cubic  feet  per  minute. 

&  =  area  of  connecting  duct,  square  feet. 

At  =  static  suction  at  throat  of  hood,  inches  of  water. 

/  =  orifice  or  restriction  coefficient,  which  varies  from  0.6  to  0.9  depending  on  the 
shape  of  the  hood. 

An  average  value  of /is  0.71,  although  for  a  well-shaped  opening  a  value 
of  0.8  may  be  used.  If  it  is  assumed  that  the  entrance  loss  of  a  hood  is 
proportional  to  the  velocity  head,  /  can  be  determined  by  the  relation: 


where 


the  velocity  head. 
the  entrance  loss. 


For  duct  ends  and  abrupt  openings  h^  =  h?  and  for  flared  openings 
&e  -  0.5AV. 

The  term  static  suction  is  not  a  good  measure  of  the  effectiveness  of  a 
hood  unless  the  area  of  the  opening  and  the  location  of  the  operation  with 
respect  to  the  hood  are  known.  This  is  clearly  indicated  by  Equation  1 
which  shows  that  the  velocity  function  at  any  point  along  the  axis  varies 
directly  as  the  area  of  the  opening  and  inversely  as  the  square  of  the 
distance.  However,  this  formula  coupled  with  Equation  2  should  serve 
to  indicate  the  velocity  conditions  to  be  expected  when  operations  are 
conducted  external  to  the  hood  opening, 

Large  Open  Hoods 

Large  hoods,  such  as  are  used  for  electroplating  and  pickling  tanks, 
should  be  subdivided  so  the  area  of  the  connecting  duct  is  not  less  than 
one-fifteenth  of  the  open  area  of  the  hood.  Frequently,  it  will  be  found 
necessary  to  branch  the  main  duct  in  order  to  obtain  a  uniform  distri- 
bution of  flow.  Canopy  hoods  should  extend  6  in.  laterally  from  the  tank 
for  every  12-in.  elevation.  In  most  cases,  hoods  of  this  type  take  advan- 
tage of  the  natural  tendency  of  the  vapors  to  rise,  and  air  velocities  may 
be  kept  low.  Cross  drafts  from  open  doors  or  windows  disturb  the  rise  of 

350 


CHAPTER  21 — INDUSTRIAL  EXHAUST  SYSTEMS' 


the  vapors  and  therefore  provision  must  be  made  for  them.  The  air 
velocities  required  also  depend  upon  the  character  of  the  vapors  given  off, 
cyanide  fumes,  for  example,  requiring  an  air  velocity  of  approximately 
75  fpm  on  the  surface  of  the  tank  and  acid  and  steam  vapors  requiring 
velocities  as  low  as  25  to  50  fpm.  The  tota.1  volume  of  air  flow  necessary 
to  obtain  these  velocities  may  be  approximately  determined  from  the 
following  simple  formula: 

Q  =  1.4PDV  (4) 

where 

Q  =  total  volume  of  air  handled  by  hood,  cfm. 

P  =  perimeter  of  the  tank,  feet. 

D  =  distance  between  tank  and  hood  opening,  feet. 

V  —  air  velocity  desired  along  edges  and  surface  of  tank,  fpm. 

Spray  Booths 

In  the  design  of  an  efficient  spray  booth,  it  is  essential  to  maintain  an 
even  distribution  of  air  flow  through  the  opening  and  about  the  object 
being  sprayed.  While  in  many  instances  spraying  operations  can  be 
performed  mechanically  in  wholly  enclosed  booths,  the  volatile  vapors 
may  reach  injurious  or  explosive  concentrations.  At  all  times  the  con- 
centrations of  these  vapors,  and  particularly  those  containing  benzol, 
should  be  kept  below  100  parts  per  million.  Spray  booth  vapors  are 
dangerous  to  the  health  of  the  worker  and  care  should  be  taken  to  mini- 
mize exposure  to  them. 

It  is  recommended  in  the  design  of  spray  booths  that  the  exhaust  duct 
be  located  in  a  horizontal  position  slightly  above  the  object  sprayed. 
Stagnant  regions  within  the  booth  should  be  carefully  avoided  or  should 
be  provided  with  a  vertical  exhaust.  The  air  volume  should  be  sufficient 
to  maintain  a  velocity  of  150  to  200  fpm  over  the  open  area  of  the  booth 
and  the  vapors  should  be  discharged  through  a  suitable  stack  to  permit 
dilution5. 

Hoods  for  Chemical  Laboratories 

Hoods  used  in  chemical  laboratories  are  generally  provided  with 
sliding  windows  which  permit  positive  control  of  the  fumes  and  vapors 
evolved  by  the  apparatus.  Their  design  should  offer  easy  access  for  the 
installation  of  chemical  equipment  and  should  be  well  lighted.  Air 
velocities  should  exceed  50  fpm  when  the  window  is  opened  to  its  maxi- 
mum height. 

DESIGN  OF  DUCT  SYSTEMS 

The  duct  system  should  be  large  enough  to  transport  the  fumes  or 
material  without  causing  serious  obstruction  to  the  air  flow.  It  is  good 
practice  to  proportion  the  ducts  to  obtain  the  desired  velocities  and 
suction  pressures  at  the  hoods,  although  in  many  cases  only  an  approxi- 
mation to  an  ideal  design  is  possible.  Many  exhaust  hoods,  and  par- 


*Far  a  discussion  of  spray  booths,  see  Special  Bulletin  No,  16,  Spray  Painting  in  Pennsylvania,  Depart- 
nwa-it  of  Labor  and  Industry,  1926,  HarrMmrg,  Pa. 

351 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

ticularly  those  used  in  buffing  and  polishing,  are  connected  by  short 
branch  pipes  to  the  main  duct  which  renders  proportioning  impractical. 

Construction 

The  ducts  leading  from  the  hoods  to  the  exhaust  fan  should  be  con- 
structed of  sheet  metal  not  lighter  than  is  shown  in  Table  4.  The  piping 
should  be  free  from  dents,  fins  and  projections  on  which  refuse  might 
catch. 

All  permanent  circular  joints  should  be  lap-jointed,  riveted  and  sol- 
dered, and  all  longitudinal  joints  either  grooved  and  locked  or  riveted 
and  soldered.  Circular  laps  should  be  in  the  direction  of  the  flow,  and 
piping  installed  out-of-doors  should  not  have  the  longitudinal  laps  at  the 

TABLE  4.    GAGE  OF  SHEET  METAL  TO  BE  USED  FOR  VARIOUS  DUCT  DIAMETERS 


DIAMETER  OP  DUCT 

GAGE  OP  MSTTAL 

8  in.  or  less  „  

24 

9  to  18  in 

22 

19  to  25  in.     _     .            

20 

26  in.  or  more  

18 

bottom.  Every  change  in  pipe  size  should  be  made  with  an  eccentric 
taper  flat  on  the  bottom,  the  taper  to  be  at  least  5  in.  long  for  each  inch 
change  in  diameter.  All  pipes  passing  through  roofs  should  be  equipped 
with  collars  so  arranged  as  to  prevent  water  leaking  into  the  building. 

The  main  trunks  and  branch  pipes  should  be  as  short  and  straight  as 
possible,  strongly  supported,  and  with  the  dead  ends  capped  to  permit 
inspection  and  cleaning.  All  branch  pipes  should  join  the  main  at  an 
acute  angle,  the  junction  being  at  the  side  or  top  and  never  at  the  bottom 
of  the  main.  Branch  pipes  should  not  join  the  main  pipes  at  points  where 
the  material  from  one  branch  would  tend  to  enter  the  branch  on  the 
opposite  side  of  the  main. 

Cleanout  openings  having  suitable  covers  should  be  placed  in  the  main 
and  branch  pipes  so  that  every  part  of  the  system  can  be  easily  reached  in 
case  the  system  clogs.  Either  a  large  cleanout  door  should  be  placed 
in  the  main  suction  pipe  near  the  fan  inlet,  or  a  detachable  section  of 
pipe,  held  in  place  by  lug  bands,  may  be  provided. 

Elbows  should  be  made  at  least  two  gages  heavier  than  straight  pipe 
of  the  same  diameter,  the  better  to  enable  them  to  withstand  the  addi- 
tional wear  caused  by  changing  the  direction  of  flow.  They  should  pref- 
erably have  a  throat  radius  of  at  least  one  and  one-half  times  the  diameter 
of  the  pipe. 

Every  pipe  should  be  kept  open  and  unobstructed  throughout  its  entire 
length,  and  no  fixed  screen  should  be  placed  in  it,  although  the  use  of 
a  trap  at  the  junction  of  the  hood  and  branch  pipe  is  permissible,  provided 
it  is  not  allowed  to  fill  up  completely. 

The  passing  of  pipes  through  fire-walls  should  be  avoided  wherever 
possible,  and  sweep-up  connections  should  be  so  arranged  that  foreign 
material  cannot  be  easily  introduced  into  them. 

At  the  point  of  entrance  of  a  branch  pipe  with  the  main  duct,  there 

259 


CHAPTER  21 — INDUSTRIAL  EXHAUST  SYSTEMS 


should  be  an  increase  in  the  latter  equal  to  their  sum.  Some  state  codes 
specify  that  the  combined  area  be  increased  by  25  per  cent.  While  this 
is  not  always  necessary  and  is  frequently  done  at  the  expense  of  a  reduced 
air  velocity,  it  is  none  the  less  advisable  where  future  expansion  of  the 
exhaust  system  is  contemplated. 

TABLE  5.    AIR  SPEEDS  IN  DUCTS  NECESSARY  TO  CONVEY  VARIOUS  MATERIALS 

MATERIAL  Am  VELOCITIES 

(FPM) 

Grain  dust _„!  2000 

Wood  chips  and  shavings _ _ •  3000 

Sawdust i  2000 

Jute  dust _ _ „ ;  2000 

Rubber  dust. - _ I  2000 

Lint. 1500 

Metal  dust  (grindings) 2200 

Lead  dusts „ j  5000 

Brass  turnings  (fine) I  4000 


Fine  coal 


4000 


Air  Velocities  in  Ducts 

When  the  static  suction  has  been  fixed  for  a  given  hood,  the  air  velocity 
in  the  duct  may  be  determined  from  Equation  2.  Air  velocities  for 
conveying  a  material  should  be  moderate.  Table  5  gives  the  velocities 
generally  employed  for  conveying  various  substances.  Equations  5a  and  5b 
may  be  used  as  tests  to  determine  the  conveying  efficiency  of  a  system6. 
Velocities  determined  from  these  formulae  should  be  increased  by  at  least 
25  per  cent  since  they  represent  the  minimum  at  which  a  stated  size  and 
density  of  material  can  be  transported. 

For  vertical  ducts:  V  =  13,300  y^y  d*-™  (5a) 

For  horizontal  ducts:  V  =  6000  y—y  <#»•*»  (5b) 

where 

V  =  air  velocity  in  duct,  feet  per  minute. 
5  —  specific  gravity  of  particles. 
d  =  average  diameter  of  largest  particles  conveyed,  inches. 

Example  2.  Granular  material,  the  largest  size  of  which  is  approximately  0.37  in.  in 
diameter,  with  a  specific  gravity  of  1.40  is  to  be  conveyed  in  a  vertical  pipe  the  velocity 
of  the  air  in  which  is  4100  fpm;  find  whether  the  material  can  be  transported  at  this 
velocity. 

Substitute  data  in  Equation  5a  and  multiply  by  1.25: 

V  =  1.25  X  13,300  X  ~|  X  0.37'-*7 

Antilog  (0.57  X  log  0.37)  =  0.568;  the  required  velocity  is,  therefore,  5500  fpm. 
Hence,  the  duct  velocity  must  be  increased  either  by  speeding  up  the  fan  or  decreasing 
th«  diameter  of  the  duct,  or  both. 


*DaHaValleF  J.  M.:  Determining  Minimum  Air  Velocities  for  Exhaust  Systems.    (A.S.H.V.E.  Journal 
Section,  Heating,  Piping  and  Air  Conditioning,  September,  1932). 

353 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

Duct  Resistance 

The  resistance  to  flow  in  any  galvanized  duct  riveted  and  soldered  at 
the  joints  may  be  obtained  from  Fig.  3,  Chapter  20.  The  pressure  drop 
through  elbows  depends  upon  the  radius  of  the  bend.  For  elbows  whose 
centerline  radii  vary  from  50  to  300  per  cent  of  pipe  diameter,  the  loss  may 
be  estimated  from  Table  6.  It  is  sometimes  convenient  to  express  the 
resistance  of  an  elbow  in  terms  of  an  equivalent  length  of  duct  of  the  same 
diameter.  Thus  with  a  throat  radius  equal  to  the  pipe  diameter  the 
resistance  is  equivalent  to  a  section  of  straight  pipe  approximately  10 
diameters  long,  while  with  a  throat  diameter  radius  lJ/£  times  the  dia- 
meter, the  resistance  is  about  the  same  as  that  of  seven  diameters  of 
straight  pipe. 

COLLECTORS 

The  most  common  method  of  separating  the  dust  and  other  materials 
from  the  air  is  to  pass  the  mixture  through  a  centrifugal  or  cyclone 
collector.  In  this  type  of  collector  the  mixture  of  the  air  and  material 
is  introduced  on  a  tangent,  near  the  cylindrical  top  of  the  collector,  and 
the  whirling  motion  sets  up  a  centrifugal  action  causing  the  compara- 
tively heavy  materials  suspended  in  the  air  to  be  thrown  against  the  side 
of  the  separator,  from  which  position  they  spiral  down  to  the  tail  piece, 
while  the  air  escapes  through  the  stack  at  the  center  of  the  collector. 

The  diameter  of  the  cyclone  should  be  at  least  3}^  times  the  diameter 
of  the  fan  discharge  duct.  When  two  or  more  separate  ducts  enter  a 
cyclone,  gates  should  be  provided  to  prevent  any  back  draft  through  a 
system  which  may  not  be  operating.  Cyclones  working  in  conjunction 
with  two  or  more  fans  should  be  designed  to  operate  efficiently  at  two- 
thirds  capacity  rating.  The  following  formula  is  useful  in  computing  the 
loss  through  a  cyclone  when  the  velocity  of  the  air  in  the  fan  discharge 
duct  is  known  : 


where 

#c  =  the  pressure  drop  through  the  cyclone,  inches  of  water. 
V  =  the  air  velocity  in  the  fan  discharge  duct,  feet  per  minute. 

If  a  cyclone  is  used  to  collect  light  dusts  such  as  buffing  wheel  dusts, 
feathers  and  lint,  the  exhaust  vent  should  be  large  enough  to  permit  an 
air  velocity  of  200  to  500  fpm.  This  will,  of  course,  require  a  cyclone  of 
larger  dimensions  than  given  for  the  foregoing  general  case. 

When  a  high  collection  efficiency  is  desired,  or  the  material  is  very  fine, 
multicyclones  may  be  used,  These  are  merely  small  cyclones  arranged  in 
parallel  which  utilize  the  principle  of  high  centrifugal  velocity  to  attain 
separation.  The  capacities  and  characteristics  of  this  type  of  separator 
should  be  obtained  from  the  manufacturers. 

Cfot-h  Filters 

Filter  cloths  are  used  when  the  material  collected  by  an  exhaust  system 
is  valuable  or  cannot  be  separated  from  the  air  with  an  ordinary  cyclone, 


CHAPTER  21 — INDUSTRIAL  EXHAUST  SYSTEMS 


They  are  also  employed  when  it  is  desirable  to  recircuiate  the  air  drawn 
from  a  room  by  the  exhaust  system,  which  otherwise  might  entail  con- 
siderable loss  in  heat.  Bag  niters  which  are  properly  housed  may  be 
operated  under  suction.  Bag  houses  used  in  the  manufacture  of  zinc  oxide 
and  other  chemical  products  are  operated  on  the  positive  side  of  the  fan. 
Wool,  cotton  and  asbestos  cloths  are  commonly  used  as  filtering 
mediums.  When  woolen  cloths  are  employed,  the  filtering  capacities  vary 
from  }/2  to  10  cfm  per  square  foot  of  filtering  surface,  depending  on  the 
character  of  the  material  collected.  The  rates  for  cotton  and  asbestos 
cloths  are  slightly  lower.  The  type  of  filter  cloth  and  the  rates  of  filtration 
depend,  of  course,  on  the  material  to  be  collected  and  the  fan  capacity. 
The  time  increase  of  resistance  varies  with  the  amount  of  material  per- 
mitted to  build  up  on  the  surface  of  the  filter  and  can  be  determined  only 
by  experiment.  The  limits  of  the  increase  may  be  regulated  by  adjust- 
ment of  the  shaking  or  cleaning  mechanism.  These  limits  may  be 
regulated  further  according  to  the  capacity  of  the  fan  and  the  effective 
performance  of  the  hoods  and  the  duct  system. 

RESISTANCE  OF  SYSTEM 

The  maintained  resistance  of  the  exhaust  system  is  composed  of  three 
factors:  (1)  loss  through  the  hoods,  (2)  collector  drop,  and  (3)  friction 
drop  in  the  pipes. 

The  loss  through  the  hoods  is  usually  assumed  to  be  equal  to  the  suction 
maintained  at  the  hoods.  The  collector  drop  in  inches  of  water  is  given 
approximately  by  Equation  6,  but  where  possible  the  resistance  of  the 
particular  collector  to  be  used  should  be  ascertained  from  the  manu- 
facturer. 

Friction  drop  in  the  pipes  must  be  computed  for  each  section  where 
there  is  a  change  in  area  or  in  velocity.  Find  the  velocities  in  each  section 
of  pipe  starting  with  the  branch  most  remote  from  the  fan.  The  friction 
drop  for  these  sections  can  be  determined  by  reference  to  Table  6.  Total 
friction  loss  in  the  piping  system  is  the  friction  drop  in  the  most  remote 
branch  plus  the  drop  in  the  various  sections  of  the  main,  plus  the  drop 
in  the  discharge  pipe. 

SELECTION  OF  FANS  AND  MOTORS 

Manufacturers  generally  provide  special  fans  for  the  collection  of 
various  industrial  wastes.  These  are  available  for  the  collection  of  coal 
dust,  wood  shavings,  wool,  cotton  and  many  other  substances.  For 
particular  features  concerning  special  fans,  consult  the  Catalog  Data 
Section  of  THE  GUIDE  and  manufacturers*  data.  When  substances 
having  an  abrasive  character  are  conveyed^  the  fan  blades  and  housing 
should  be  protected  from  wear.  This  may  be  accomplished  by  placing  a 
collector  on  the  negative  side  of  the  fan  or  by  lining  the  housing  and 
blades  with  rubber. 

If  no  future  expansion  of  an  exhaust  system  is  contemplated,  the  fart 
motor  should  be  chosen  to  provide  the  calculated  air  volume.  Should, 
however,  the  exhaust  system  be  required  to  handle  more  air  in  the 
future,  the  motor  should  be  adequate  for  the  maximum  load  anticipated.. 

355 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

Further  information  regarding  the  choice  of  fans  and  motors  is  given  in 
Chapter  17. 

PROTECTION  AGAINST  CORROSION 

The  removal  of  gases  and  fumes  in  many  chemical  plants  requires  that 
metals  used  in  the  construction  of  the  exhaust  system  be  resistant  to 

TABLE  6.    Loss  THROUGH  90-DEG  ELBOWS 


ELBOW  CENTRE  LENH  RADIUS  IN  PBB  CENT 
or  PIPE  DIA,MZTEB 

Loss  n?  PEE  CENT  OF  VELOCITY  HEAD 

50 
100 
150 
200  to  300 

75 
26 
17 
14 

chemical  corrosion.  A  list  of  the  materials  which  may  be  used  to  resist 
the  action  of  certain  fumes  is  given  in  Table  7.  Hoods  and  ducts  when 
short,  may  frequently  be  constructed  of  wood  and  be  quite  effective, 

TABLE  7.    MATERIALS  TO  BE  USED  FOR  THE  PROTECTION  OF 
EXHAUST  SYSTEMS  AGAINST  CORROSION 


TTPB  OF  FUME  COJTVETBD 

PROTECTIVE  MATERIAL  TO  BE  USED 

Chlorine  .  
Hydrogen  sulphide  
Ammonia.  -  

Rubber  lining  or  chrome-nickel  alloys 
Aluminum  coated  iron,  aluminum,  high  chrome-nickel  alloys 
Iron  or  steel 

Sulphurous  gases  
Hydrochloric  acid-  

Nitrous  gases 

High  chrome-nickel  alloys 
Rubber  lining,  chrome-nickel  alloys 
Nickel-chrome  alloys 

^Condensed  from  data  given  by  Chilton  and  Huey  (Industrial  and  Engineering  Chemistry,  Vol.  24, 1932). 

Rubberized  paints  are  available  and  may  be  applied  as  protective  coatings 
in  handling  such  gases  and  fumes  as  chlorine  and  hydrochloric  acid. 


PROBLEMS  IN  PRACTICE 

1  •  Should  individual  operations  be  served  by  an  individualized  dust  collector 
system? 

Yes,  if  operations  are  usually  kept  individual  in  a  group  of  machines. 

2  •  Axe  state  regulatory  requirements  as  to  suction  applicable  to  all  sorts  of 
dust  collecting  installations? 

As  a  rule  the  regulations  refer  only  to  grinding  wheel  and  buffing  wheel  systems.    They 
are  needed  for  many  other  industrial  processes. 


CHAPTER  21 — INDUSTRIAL  EXHAUST  SYSTEMS 


3  •  What  is  the  most  common  method  of  reducing  total  air  yolumes  handled 
in  cases  employing  large  hoods  over  apparatus  covering  a  large  area? 

The  use  of  the  petticoat  or  double  hood  which  permits  a  comparatively  high  air  velocity 
at  the  rim  of  the  hood  and  controllably  small  velocities  in  the  center. 

4  •  What  other  types  of  collectors  are  available  for  use  in  the  place  of  cyclones 
and  niters  when  chemical  and  physical  conditions  obviate  the  possibility  of  the 
use  of  them? 

Devices  such  as  scrubbers  and  contactors,  using  water  or  other  contacting  liquids, 
electrical  precipitators,  and  dynamical  precipitators. 

5  •  What  is  the  most  frequent  error  made  in  dust  collector  system  design? 

The  omission  of  some  means  of  putting  into  the  workroom  air  having  the  proper  charac- 
teristics to  replace  that  which  has  been  exhausted. 

6  •  Are  there  available  means  for  testing  the  performance  of  dust  collecting 
systems  when  they  are  required  to  meet  high  industrial  hygienic  standards? 

Yes.  Such  means  are  set  up  by  the  United  States  Public  Health  Service  and  by  the 
Standard  Code  for  Testing  Centrifugal  Fans  (Chapter  41). 

7  •  Why  is  it  not  permissible  to  connect  up  emery  wheels  and  buffing  wheels  to 
the  same  exhaust  system? 

Emery  wheels  and  buffing  wheels  should  be  handled  by  separate  systems  because  of  the 
fire  hazard,  as  it  is  possible  for  sparks  from  the  emery  wheels  to  ignite  the  lint  and  dust 
from  the  buffing  wheels  when  both  are  carried  through  the  same  system. 

3  •  Give  an  important  characteristic  of  centrifugal  type  dust  collectors  which 
should  be  given  consideration  when  applying  this  type  of  collector  to  instal- 
lations requiring  high  separating  efficiencies. 

The  separating  action  of  a  cyclone  or  centrifugal  type  collector  depends  largely  on 
centrifugal  force.  Reducing  the  radius  of  air  flow  increases  the  centrifugal  force  for  a 
given  velocity  of  flow.  Accordingly,  the  smaller  size  units  usually  give  higher  separating 
factors,  and  better  results  can  sometimes  be  obtained  by  using  a  number  of  small  col- 
lectors instead  of  one  large  unit. 

9  •  Mention  some  general  suggestions  relating  to  the  design  of  efficient  in- 
dustrial exhaust  systems. 

a.  Endeavor  to  obtain  a  maximum  degree  of  effectiveness  with  a  minimum  volume  of  air, 
by  the  use  of  well  designed  hoods  closing  in  the  sources  of  fumes  or  material  to  be  removed 
so  located  as  to  take  advantage  of  the  natural  direction  taken  by  the  fumes  or  materials 
when  leaving  their  source. 

b.  Give  particular  care  to  the  velocity  of  flow.    The  duct  velocities  for  material  con- 
veying systems  must  be  high  enough  to  properly  carry  the  material,  but  they  should  not 
be  higher  than  necessary  because  excessive  velocities  increase  the  pressure  requirements 
and  result  in  a  waste  of  power. 

c.  Select  the  type  of  fan  best  suited  to  the  job.    For  installations  where  stringy  material 
is  handled  do  not  use  a  fan  wheel  which  has  a  shroud. 

J.  When  handling  the  refuse  from  various  machines,  study  the  grouping  and  operating 
cycles  of  the  machines.  Connecting  a  large  number  of  machines  into  one  system  is 
frequently  very  uneconomical. 

e.  Avoid  unnecessary  distances  and  bends  in  laying  out  the  piping  system. 

10  •  The  static  pressure  measured  at  the  throat  of  a  buffing  wheel  hood  is  2  in. 
and  the  velocity  head  measured  with  a  Pi  tot  tube  is  1.6  in.     Calculate  the 
restriction  coefficient  f. 

357 


AMERICAN  SOCIETY  of.  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

From  Equation  2,  V  =  4005  /  V~ht- 

From  the  theory  of  air  flow,  V  =  4005  \/  hv. 

Hence,  \/Tv  -  / 


1.1  •  A  tank}  4  ft  by  8  ft,  contains  a  fluid  which  gives  off  injurious  vapors.  A 
large  hood  is  located  30  in.  above  the  top  of  tfce  tank  and  extends  slightly  over 
its  edges.  Assuming  that  a  velocity  of  60  fpm  is  required  to  adequately  control 
the  vapors  near  the  edges  of  the  tank,  calculate  the  air  flow  required. 

Using  Equation  4,  P ;  =  2  X  4  -f  2  X  8  =  24  ft;  D  =  30  inches  =  2.5  ft;  V  =  60  fpjn. 
Hence,  Q  =  1.4  X  24  X  2.5  X  60  -  5.040  cfm. 

12  •  Silica  dust  with  a  specific  gravity  of  2.65  is  being  conveyed  in  a  duct  system; 
The  velocity  measured  in  a  vertical  portion  of  the  system  is  found  to  be  2700 
fpm.  What  is  the  maximum  diameter  particle  transported  at  this  velocity? 

Using  Equation  5a,  2700  =*  13,300  X  ~~   X  ^-57° 

o.OO 

from  which 

d  »  (0.28)1-75  -  0.11  in. 


358 


Chapter  22 

FAN  SYSTEMS  OF  HEATING 

Types  of  Systems,  Blow -Through,  Draw-Through,  Heating  Units, 
Design,  Temperatures,  Weight  of  Air  to  be  Circulated,  Tempera- 
ture Loss  in  Ducts,  Heat  Supplied  Heating  Units  and  Washer, 
Grate  Area,  Boiler  Selection,  Weight  of  Condensate,  Static  Pres- 
sure, Fans  and  Control 

A  FAN  system  of  heating  depends  upon  fans  and  blowers  to  distribute 
air  through  ducts  from  one  centrally  located  plant.  This  chapter 
considers  heating  and  humidifying  systems  of  this  type  whereas  similar 
systems  arranged  for  cooling  and  dehumidifying  are  discussed  in  Chapter 
9.  A  special  type  of  central  fan  system,  the  mechanical  warm  air  or  fan 
furnace  system,  which  is  especially  adapted  to  residences,  churches,  halls, 
and  other  small  buildings,  is  covered  in  Chapter  23. 

TYPES  OF  SYSTEMS 

In  the  indirect  type  of  central  fan  heating  and  air  conditioning  systems, 
steam  is  usually  the  medium  by  which  heat  is  transferred  from  the  boiler, 
or  other  source  of  heat,  to  the  heating  units.  If  the  system  is  intended 
solely  for  heating,  the  air  is  passed  over  one  or  more  stacks  or  batteries  of 
heating  units  and  then  conveyed  to  the  spaces  for  which  it  is  intended 
through  a  system  of  ducts.  In  some  cases,  a  predetermined  amount  of 
outside  air  is  introduced  for  ventilating  purposes,  whereas  in  others  the 
moisture  content  is  controlled  by  passing  the  air  through  a  washer  or 
humidifier.  If  the  apparatus  is  designed  to  control  simultaneously  the 
temperature,  humidity,  air  motion,  and  distribution,  it  is  known  as  an  air 
conditioning  system. 

In  the  split  system,  the  heating  is  accomplished  by  means  of  radiators  or 
convectors,  and  the  ventilating  or  air  conditioning  by  means  of  the  central 
fan  apparatus.  In  the  combined  system,  the  entire  operation  of  heating, 
ventilating,  and  air  conditioning  is  handled  by  the  central  fan  system. 

A  common  arrangement  of  the  central  fan  system  of  heating  is  illus- 
trated by  Fig.  1  and  consists  of  a  fan,  a  heating  unit  (heater)  enclosed  by  a 
sheet  metal  casing  connected  with  the  suction  side  of  the  fan,  a  sheet1 
metal  casing  connected  to  the  heating  unit  casing  run  to  the  outside  of  the" 
building  and  provided  with  an  adjustable  opening  inside  the  building  for 
recirculation  of  the  air  when  desired,  and  a  duct  system  attached  to  the 
fan  outlet  to  convey  and  distribute  tlie  air  to  various  parts  of  the  building 
to  be  warmed  by  the  apparatus.  The  fan  is  ordinarily  motor-driven ;  there 
are,  Ifo^ever,  many  cases  when  a  direct-connected  steam  engine  may  be 
used  to  advantage.  In  this  event  the  exhaust  from  the  engine  can  be  cori- 

359 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

nected  to  one  or  more  sections  of  the  heater,  depending  upon  the  con- 
densation rate  of  the  engine.  The  recirculation  duct  connected  with  the 
opening  in  the  suction  duct  should  be  extended  to  a  point  as  near  the 
floor  as  possible. 

When  ventilation  is  not  a  requirement  or  is  considered  relatively  unim- 
portant, as  in  shop  and  factory  heating,  and  the  number  of  persons  vitiat- 
ing the  air  is  small  compared  with  the  cubical  contents  of  the  building,  or 
the  process  does  not  generate  obnoxious  gas  or  vapors,  the  air  may  be 
recirculated,  sufficient  outside  air  for  ventilation  being  supplied  by  infiltra- 


Rotting  Shutter- 


Oubid*  Wall 


By-pass  Damper 

FIG.  1.    ARRANGEMENT  OF  A  CENTRAL  FAN  HEATING  SYSTEM 
(DRAW-THROUGH) 


Canvas  Connection 


Heater 


Foundation 


V 

_ 

Supply  Duct  | 

By-pass  Damper 


Floor 


FIG.  2.    ARRANGEMENT  FOR  HEATING  UNIT  (BLOW-THROUGH) 

tion.   The  amount  of  heat  to  be  supplied  the  heating  unit  in  this  case  is  the 
same  as  would  be  required  for  a  direct  radiation  installation. 

When  ventilation  is  a  requirement  to  be  met,  an  arrangement  similar  to 
that  shown  by  Fig.  1  may  be  employed.  Since  the  amount  of  air  necessary 
for  heating  is  generally  in  excess  of  the  amount  required  for  ventilation, 
considerable  fuel  economy  may  be  effected  by  recirculating  a  portion  of 
the  air.  In  this  case  only  sufficient  outside  air  is  drawn  into  the  system  to 
meet  the  ventilation  requirement  and  the  remainder  of  the  air,  required 
for  heating,  is  recirculated.  This  may  be  readily  effected  by  an  arrange- 
ment of  ducts  and  dampers  on  the  suction  side  of  the  fan  as  previously 
mentioned.  If  the  outside  air  introduced  is  to  be  washed  or  conditioned 
the  washer  or  humidifier  and  tempering  coil  may  be  added  between  the 
inlet  for  the  recirculated  air  and  the  fresh  air  intake. 

360 


CHAPTER  22 — FAN  SYSTEMS  OF  HEATING 


Blow-Through,  Draw-Through 

When  the  heating  unit  is  located  on  the  suction  side  of  the  fan,  the 
system  is  known  as  draw-through.  (See  Fig.  1.)  When  the  heating  unit 
is  located  in  the  discharge  from  the  fan,  the  system  is  known  as  blow- 
through.  (See  Fig.  2.)  The  draw-through  combination  is  used  for  factory 
and  toilet  room  installations  because  a  more  compact  arrangement  of 
the  apparatus  usually  is  possible.  In  addition,  air  leakage  will  be  inward. 
The  blow-through  combination  is  used  principally  in  schools  and  public 
buildings,  and  for  all  booster  coil  arrangements  where  different  tempera- 
tures and  independent  temperature  regulation  are  required  for  different 
heated  spaces.  In  public  building  installations,  the  fan  frequently  blows 
the  heated  air  into  a  plenum  chamber  from  which  the  air  ducts  radiate  to 
the  various  rooms  of  the  building;  this  arrangement  is  sometimes  called 
the  plenum  system. 

HEATING  UNITS 

The  heating  units  for  central  fan  systems  using  steam  as  the  heating 
medium  may  be  classified  as  (1)  tempering  coils,  (2)  preheater  coils,  (3) 
reheater  coils,  (4)  booster  coils,  and  (5)  water  heaters,  either  open  or 
closed.  Tempering  coils  are  used  with  ventilating  and  air  conditioning 
systems  for  raising  the  temperature  of  the  outside  cold  air  to  above  freez- 
ing, or  32  F.  They  are  not  required  for  heating  systems  where  all  of  the 
air  is  recirculated,  since  the  temperature  of  the  recirculated  air  will  be 
above  freezing.  Preheater  coils  are  used  with  air  conditioning  systems  to 
raise  the  temperature  of  the  air  from  that  leaving  the  tempering  coils  to 
such  a  temperature  that  in  passing  through  the  water  sprays  of  the  washer 
(without  water  heater)  the  air  will  become  partially  saturated  (adia- 
batically)  having  a  moisture  content  corresponding  to  the  required  dew- 
point  temperature.  Preheater  coils  therefore  supply  heat  as  necessary  to 
control  the  dew-point  temperature.  The  reheater  coils  are  used  to  raise  the 
temperature  of  the  air  leaving  the  tempering  coils  (in  the  case  of  a  heating 
or  ventilating  system)  or  the  air  leaving  the  washer  (in  the  case  of  an  air 
conditioning  system)  to  that  necessary  to  maintain  the  desired  tempera- 
ture in  the  rooms  or  spaces  to  be  heated  or  conditioned,  except  where 
booster  coils  are  used,  in  which  case  the  reheater  coils  raise  the  air  tem- 
perature to  approximately  room  temperature,  or  slightly  higher.  Booster 
coils  are  installed  in  the  duct  branches  to  control  the  temperature  of  the 
air  entering  the  rooms  or  spaces  for  which  it  is  intended.  Water  heaters  are 
used  on  an  air  conditioning  system  to  control  the  dew-point  temperature. 
They  are  used  mainly  for  industrial  work,  seldom  for  comfort  conditioning. 
They  are  not  used  where  preheater  coils  are  employed.  The  open  type 
supplies  steam  directly  to  the  spray  water,  while  the  closed  type  utilizes  a 
heat  interchanger  by  which  the  steam  imparts  its  heat  to  the  spray  water. 
Where  water  heaters  are  required  for  comfort  conditioning,  the  closed 
type  is  used. 

The  heating  units  for  central  fan  systems  in  use  at  the  present  time  con- 
sist either  of  pipe  coils,  finned  tubes  of  steel,  copper,  brass  or  other  metal, 
cast-iron  sections  with  extended  surfaces,  or  the  cellular  type.  Steam  is 
passed  through  these  heating  units  and  the  air  to  be  heated  is  passed  over 
their  exterior  surfaces. 

361 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

In  selecting  a  heating  unit  for  any  particular  service,  the  choice  should 
be  based  on  the  desired  requirements  as  follows: 

1.  Final  temperature  desired. 

2.  Loss  in  pressure  for  friction)  of  air  passing  over  the  heating  unit. 

3.  Air  velocity  over  the  heating  unit. 

4.  Free  area  or  face  area  of  heating  unit. 

5.  Ratio  of  heating  surface  to  net  free  (or  face)  area. 

6.  Air  volume  required. 

7.  Number  of  rows  of  pipes,  tubes,  or  sections. 

8.  Amount  of  heating  surface. 

9.  Steam  pressure  drop  through  the  heating  unit. 
10.  Weight  of  heating  unit. 

Final  Temperature  Desired.  The  choice  of  a  heating  unit  is  Jargely 
influenced  by  the  final  temperature  desired,  when  the  entering  air  tem- 
perature and  steam  pressure  available  at  the  heating  unit  are  specified. 
These  data  are  obtainable  from  manufacturers'  catalogs. 

Loss  in  Air  Pressure  (or  Friction).  The  allowable  friction  through  the 
heating  unit  is  one  of  the  first  factors  to  be  determined  in  the  selection  of 
the  apparatus.  The  velocities  of  air  through  various  types  of  heating 
units  will  not  necessarily  be  the  same,  but  for  any  particular  job  the 
velocity  through  the  heating  unit  should  be  a  secondary  consideration  and 
the  allowable  friction  or  air  pressure  loss  should  be  fixed  approximately 
before  proceeding  with  the  selection  of  the  heating  unit.  The  loss  in  air 
pressure  (or  friction)  through  the  heating  unit  should  not  exceed  a  pre- 
determined maximum  allowable  amount  for  economical  operation  and  for 
moderate  size  and  first  cost  of  installation. 

In  public  building  work,  the  maximum  allowable  friction  through  both 
tempering  coil  and  reheater  coils  should  never  exceed  ^  in.  of  water  and 
it  is  advisable  that  the  friction  be  kept  considerably  lower  than  this  figure 
if  possible.  A  tempering  coil  friction  ranging  from  0.10  to  0.20  in.  of  water 
is  considered  satisfactory.  The  air  pressure  loss  for  reheaters  ordinarily 
ranges  from  0.20  to  0.40  in.  of  water.  In  factory  work,  the  maximum 
friction  through  the  heater  should  never  exceed  0.8  in.  or  1  in.  of  water 
and  it  is  advisable  to  figure  the  heaters  at  lower  frictions  if  possible. 

Velocity  through  Heating  Unit.  This  velocity  has  generally  been  given 
in  manufacturers*  tables  as  being  measured  at  70  F  and  in  most  cases 
refers  to  the  velocity  through  the  net  free  area  of  the  heating  unit,  or 
through  the  net  space  between  the  pipes,  tubes  or  sections.  Although 
most  manufacturers  give  suitable  velocities  measured  at  70  F,  certain 
manufacturers  show  velocities  measured  at  65  F  and  others  indicate 
velocities  measured  at  the  average  air  temperature  through  the  heating 
unit.  Many  new  heating  units,  however,  specify  net  face  areas  with  cor- 
responding velocities  instead  of  velocities  through  net  free  areas.  In 
either  case,  manufacturers  publish  the  corresponding  friction  or  air- 
pressure  loss  in  tables.  The  velocity  through  the  net  free  area  of  the 
heating  unit  averages  about  1000  fpm  and  that  through  the  net  face  area 
about  500  fpm. 

The  volume  of  air  to  be  heated  in  any  particular  case  is  determined  after 
consideration  of  the  ventilation  requirements,  heat  losses,  and  quantity  of 
air  required  for  proper  circulation,  as  explained  in  Chapters  2  and  7. 

362 


CHAPTER  22 — FAN  SYSTEMS  OF  HEATING 


The  number  of  rows  of  pipes,  tubes,  or  sections  or  the  amount  of  heating 
surface  to  be  used  may  be  selected  from  manufacturers'  catalogs  after  the 
quantity  of  air  handled  and  the  heat  load  are  known.  Savings  in  oper- 
ating expense  or  cost  of  installation  should  result  from  a  proper  selection 
of  heater  and  by-pass  areas.  For  example,  instead  of  having  the  entire 
air  quantity  go  through  a  one-row  heating  unit,  it  may  be  advantageous 
to  use  a  two-row  heating  unit  and  a  properly  sized  by-pass.  Thus,  when 
no  heating  is  being  done,  a  suitable  by-pass  damper  may  be  opened  to 
place  a  lighter  load  on  the  fan. 

The  steam  pressure  drop  through  the  heating  unit  is  also  tabulated  in 
manufacturers*  data  tables.  The  sizing  of  steam  supply  and  return 
piping,  allowing  for  drops  through  heating  units,  is  explained  in  Chapter 
32. 

Weight  of  Heating  Unit.  In  the  design  of  a  heating  system,  the  weight 
limitations  of  heating  units  are  determined  by  the  location  of  the  units. 
Obviously,  if  there  is  no  loading  limitation  imposed,  any  type  of  heating 
unit  may  be  selected.  On  the  other  hand  if  the  heating  unit  is  to  be  hung 
from  the  ceiling,  it  may  be  desirable  to  use  the  lightest  unit  which  will 
accomplish  the  work  required. 

DESIGNING  THE  SYSTEM 

The  general  procedure  for  the  design  of  central  fan  systems  is  as 
follows : 

1.  Calculate  the  heat  loss  for  each  room  or  space  to  be  heated. 

2.  Determine  volume  of  outside  air  to  be  introduced, 

3.  Assume  or  calculate  temperature  of  air  leaving  registers  or  supply  outlets. 

4.  Calculate  weight  of  air  to  be  circulated. 

5.  Estimate  temperature  loss  in  duct  system. 

6.  Calculate  heat  to  be  supplied  the  heating  units  and  washer. 

7.  Select  heating  units  and  washer  from  manufacturers*  data  and  performance  curves. 

8.  Calculate  total  heat  to  be  supplied, 

9.  Calculate  grate  area  and  select  boiler. 

10.  Design  duct  system. 

11.  Calculate  total  static  pressure  of  system. 

12.  Select  fan,  motor,  and  drive. 

The  heat  losses  (If)  should  be  calculated  in  accordance  with  the  pro- 
cedure outlined  in  Chapter  7.  If  a  positive  pressure  is  maintained  by  the 
central  fan  system  in  the  room  or  space  to  be  ventilated  or  conditioned, 
there  will  ordinarily  be  very  little  infiltration  of  cold  outside  air  through 
the  cracks  and  crevices  of  the  space.  Consequently,  the  volume  of  air 
introduced  into  the  space  at  the  assumed  or  calculated  outlet  temperature 
need  only  be  sufficient  to  provide  for  the  transmission  losses,  plus  about 
one-third  of  the  infiltration  losses.  The  exfiltration  of  heated  or  con- 
ditioned air  through  the  cracks  and  crevices  of  the  space  should  be  pro- 
vided for  by  making  the  usual  allowance  for  the  infiltration  losses  in 
arriving  at  the  total  heat  loss  of  the  space.  The  air  required  to  make  up 
for  this  exfiltration  of  heated  or  conditioned  air  will  be  brought  in  at  the 
outside  air  intake  and  may  be  included  as  a  part  of  the  outside  air  neces- 

363 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

sary  for  the  ventilating  requirements.  The  heat  required  to  raise  this  air 
to  the  conditions  maintained  in  the  room  must  be  provided  by  the  tem- 
pering coils,  preheater  coils,  and  reheater  coils.  If  a  positive  pressure  is 
not  maintained  in  the  room  or  space  to  be  conditioned,  the  normal  in- 
filtration of  outside  cold  air  will  take  place  in  this  room,  and  the  outlet 
temperature,  together  with  the  required  air  volume  at  this  temperature, 
must  be  sufficient  to  provide  for  both  infiltration  and  transmission  losses. 

Volume  of  Outside  Air 

The  volume  of  outside  air  required  for  ventilation  or  air  conditioning 
purposes  may  be  determined  from  data  in  Chapter  2.  In  no  case  shall 
less  than  10  cfm  per  person  be  introduced. 

The  heat  required  to  warm  the  outside  air  introduced  for  ventilation 
purposes  (Ho)  may  be  determined  by  means  of  the  following  formula: 

Ho  «  0.24  (t  -  to)  M0  (1) 

where 

0.24  =  specific  heat  of  air  at  constant  pressure. 
/  =  room  temperature,  degrees  Fahrenheit. 
to  =  outside  temperature,  degrees  Fahrenheit. 

MO  —  weight  of  outside  air  to  be  introduced  per  hour,  in  pounds  =  d0Q0. 
Qo  =  volume  of  outside  air  to  be  introduced,  cubic  feet  per  hour. 
d0  —  density  of  air  at  t0,  pounds  per  cubic  foot. 

Example  1  .  A  building  in  which  the  temperature  to  be  maintained  at  70  F  requires 
10,000  cfm.  If  the  outside  temperature  is  20  F,  how  much  heat  will  be  required  to  warm 
the  air  introduced  for  ventilation  purposes  to  the  room  temperature? 

Solution.  Qo  =  10,000  X  60  =  600,000  cfh;  d0  »  0.08276  (Table  3,  Chapter  1); 
Mo  =  0.08276  X  600,000  =  49,656  Ib;  t  =  70  F;  t0  =  20  F;  H0  «  0.24  X  (70  -  20) 
X  49,656  =  595,872  Btu  per  hour. 

Temperature  of  Air  Leaving  Registers 

If  the  system  is  to  function  only  as  a  heating  system,  that  is,  entirely  as 
a  recirculating  one,  the  temperature  of  the  air  leaving  the  register  outlets 
must  be  assumed.  For  public  buildings,  these  temperatures  may  range 
from  100  to  120  F,  whereas  for  factories  and  industrial  buildings  the  out- 
let or  register  temperature  may  be  as  high  as  140  F.  In  no  case  should  the 
outlet  temperature  exceed  these  values. 

For  ventilating  or  conditioning  systems,  the  temperature  of  the  air 
leaving  the  supply  outlets  may  be  estimated  by  means  of  the  following 
formula  : 


M  (2) 

where 

ty  =  outlet  temperature,  degrees  Fahrenheit. 

H  =  heat  loss  of  room  or  space  to  be  conditioned,  Btu  per  hour. 

Q  =  total  volume  of  air  to  be  introduced  at  the  temperature  /,  cubic  feet  per  hour. 

If  the  outlet  temperature  (ty)  as  determined  from  Equation  2  exceeds 
120  F  for  public  buildings,  or  140  F  for  factories  or  industrial  buildings, 


CHAPTER  22 — FAN  SYSTEMS  OF  HEATING 


these  respective  outlet  temperatures  should  be  used  as  factors  in  the 
following  equation  to  determine  the  volume  of  air  to  be  introduced  into 
the  room  or  space: 

_    55.2H 
Q  ~  (h  -  t)  (3) 

Example  £.  The  heat  loss  of  a  certain  auditorium  to  be  conditioned  is  100,000  Btu  per 
hour.  The  ventilating  requirements  are  90,000  cu  ft  per  hour  and  the  room  temperature 
70  F.  Determine  the  outlet  temperature. 


Solution.    Substituting  in  Formula  2, 

55.2  X  100,000 
h  90,000 


-f  70  »  131.3  F 


Inasmuch  as  this  temperature  is  excessive,  it  will  be  necessary  to  assume  an  outlet 
temperature,  which  will  be  taken  as  120  F,  and  to  calculate  the  amount  of  air  to  be 
introduced  into  the  room  at  this  temperature  to  provide  for  the  heat  loss.  Substituting 
in  Equation  3, 

Q  _  65^100^000    =  U(WOO  cfh  (at  temperature  0 

Weight  of  Air  to  be  Circulated 

The  total  weight  of  air  to  be  introduced  into  the  room  or  space  to  be 
heated  or  conditioned  (M)  is  given  by  the  following  formulae: 


M   =  Mo  -f  Mr  (5) 

Mo  =  doQo  (6) 

where 

d  =  density  of  air  at  temperature  t,  pounds  per  cubic  foot. 

do  =  density  of  air  at  temperature  /o,  pounds  per  cubic  foot. 

Qo  =  volume  of  outside  air  at  temperature  to. 
M0  =  weight  of  outside  air,  pounds. 
Mr  =  weight  of  recirculated  air,  pounds. 

Example  8.    Using  the  data  of  Example  2  and  an  outside  temperature  of  20  Ft  what 
will  be  the  values  of  M,  M0  and  Afr? 

Solution,    d  -  0.07495  ;&>  =jQ,QS276;()  =  110,400;  Q0  -  90,000;  H  =  100,000. 
_  100,000 


~  0.24  X  (120  -  70) 
MG  =  0.08276  X  90,000  -  7,448  lb 
MT  -  M  -  Mo  -  8,333  -  7,448  =  885  lb 

Temperature  Loss  in  Ducts 

The  allowances  to  be  made  for  loss  in  transit  through  the  duct  system 
(/,)  are  as  follows: 

1.  When  the  duct  system  is  located  in  the  enclosure  to  which  the  air  is  being  delivered, 
as  in  a  factory,  it  may  be  assumed  that  there  is  no  loss  between  the  r^heater  cotl  and  the 
point  or  points  of  discharge  into  the  enclosure. 

365 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

2.  For  ducts  in  outside  walls  or  attics,  or  other  exposed   places,  allow  Q.25  F  per 
linear  foot  of  uninsulated  duct. 

3.  For  ducts  run  underground  an  allowance  shall  be  made  based  on  the  estimated  heat 
loss  of  the  duct,  assuming  the  average  temperature  of  the  ground  to  be  55  F. 

Heat  Supplied  Heating  Units  and  Washer 

The  following  cases  may  arise  in  practice : 

A.  The  heating  of  the  building  is  done  entirely  by  means  of  a  central  fan  system,  all 
of  the  air  being  drawn  from  the  outside. 

B.  Similar  to  A,  except  that  all  of  the  air  is  recirculated. 

C.  A  portion  of  the  air  is  recirculated,  and  the  remainder  is  drawn  in  from  the  outside. 

D.  Air  at  the  same  temperature  is  to  be  delivered  to  all  the  rooms.    A  constant  relative 
humidity  is  maintained  in  the  building  and  all  of  the  air  circulated  is  drawn  from  outside 
the  building.    (Not  applicable  to  the  heating  of  various  rooms  where  individual  control 
of  each  room  is  desired.) 

E.  Outside  air,-  return  air,  and  by-pass  air  are  used  with  the  reheater  located  in  by- 
pass air  chamber. 

F.  Arrangement  of  apparatus  where  individual  control  of  the  temperature  for  .each 
room  is  required  in  conjunction  with  air  washer  equipment  to  maintain  a  constant 
relative  humidity  in  the  rooms.    The  airi  "washer  is  provided  with  a  water  heater  for  the 
spray  water ,:capable  of  fully  saturating  the  air.    A  section  of  preheater  may  be  used  for. 
this  purpose  in  place  of  the  water  heater.    With  this  arrangement  and  with  a  uniform 
temperature  of  air  entering  the  rooms,  it  is  impossible  to  maintain  the  same  room  tem- 
perature throughout  the  building  because  the  weight  of  air  to  be  delivered  to  each  room 
is  determined  and  fixed  by  the  ventilating  requirements. 

In  analyzing  these  cases,  the  following  symbols  will  be  used : 

H  =  heat  loss  of  the  room  or  building,  Btu  per  hour. 
Hi  «  heat  to  be  supplied  to  the  reheater  coil,  Btu  per  hour. 

Hz  =  heat  supplied  tempering  coil,  or  compering  >eoii  and  preheater*  Btu  per  hour. 
HZ  =  heat  supplied  air  washer  by  wa#er  heater,  Btii  per  hour. 
#4  =  heat  to  be  supplied  booster  coil,  Btu  per  hour. 

M  —  weight  of  air  to  be  introduced  into  the  room  or  building,  pounds  per  hour. 
'Mi  «  weight  of  recirculated  air,  pounds  per  hour, , 
Mb  — ,  weight  of  air  by-passing  washer,  pounds. per  hour. 
jlf0  =""  weight  of  air  drawn  in  from  outside,  pounds  per  hour. 

to  —  mean  temperature  of  outside  air,  degrees  Fahrenheit* 

/  =  mean  air  temperature  to  be  maintained  in  the  room  or  building,  degrees 
Fahrenheit.         .     . 

h  =  mean  temperature  of  the  air  entering  the  reheater  coil. 
/2  =  mean  temperature  of  the  air  leaving  the  reheater  coil. 
tz  =  temperature  loss  in  the  duct  system. 
ty  =  temperature  of  the  air  leaving  the  duct  outlets,,; 
tK  —  average  temperature  of  air  entering  tempering  coil. 
&#  —  temperature  of  air  entering  washer. 
0.24  =  specific  heat  of  air  at  constant  pressure1. 

366 


CHAPTER  22 — FAN  SYSTEMS  OF  HEATING 


Rolling  Shutter- 


/  Steam 
Control  Valve    ||          _  Control  Valve 


-Air  Leaving  Fan  at  ty 


Outside  Air        J 
Louvres  "*-*-r  , 


Outside  Wall 


By-pass  Damper 

FIG.  3.     HEATING  UNIT  AND  FAN  ARRANGED  FOR  OUTSIDE  AIR  CIRCULATION  (Case  A) 

Case  A .      (Fig.  3)  All  of  the  air  circulated  to  be  drawn  from  outside  the  building,  in 
which  case  tx  —  t0. 

-  *o)  M0  •      (7) 

.       .      ,(8) 


Hi  =  0.24  fe  -  id  Mo 


Example  4-  The  heat  loss  H  for  a  certain  factory  building  is  700,000  Btu  per  hour. 
The  mean  inside  temperature  t  to  be  maintained  is  65  F.  The  assumed  outside  air  tem- 
perature to  is  0  F;  tz  =  0,  ty  «  /2  and  is  assumed  to  be  140  F.  The  temperature 
leaving  the  tempering  coil  is  assumed  to  be  35  F.  Required,  Hi  and  Hi.  From  Equation  4, 


M  * 


700,000 


0.24  (140  -  65) 


38,889  Ib  per  hour. 


Hi  =  0.24  X  (35  -Q)  X  38,889  »  326,667  Btu  per  hour. 
Hi  «  0.24  X  (140  -  35)  X  38,889  =  980,003  Btu  per  hour. 
•TrJs:H~  Hi  .*  326,667  -1-  980,003  »  1,306,670  Btu  per  hour. 


'Air  Returned 
from  Heated  Space 

^T 

CH 

,Stearn 
^-Automatic  Valve  , 

-Air  Leaving  Fan  at  ty 
^Pulley 

TT      -    '  . 

/Foundation 

Heater-^ 

\r 

X 

<2 

Fan 

I 

ELEVATION 

FIG.  4,    ARRANGEMENT  FOR  RECIRCULATION  (Case  B) 

:  •  (Fig.  4)  All  of  the  air  is  to  be  recirculated,  in  which  case  t\  =  /. 

,',,          -       M*  =  38,889  Ib 

Mi  ^  0,24  (^  -  h)  Mr 
:..',„.  Hi  «  0.24  (140  -  65)  X  38,889  =  700,000  Btu  per  hour. 

This  Example  illustrates  the  saving  in  fuel  consumption  by  the  *^-^*- 
culation  of  the  air.  The  heat  to  be  supplied  the  apparatus  is  the  same  ais 
that  required  for  a  direct  system  of  heating  and  is  equal  to  the  heat  loss 
of  th^Fbuilding  '(Hi  =!H),  in  the  example  700,000  Btu  per  hour  as 
compared  with  1,306,670  for  Case  A. 

367 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


Rolling  Shutter-  £ 

/                                           /S 
;             t      Control  Valve    II 

/  Recirculated  Air                 ||  1 
'     .      yy/   .                            U  J 

earn 
Cor 

g^ 

'      i 

trol  Valve 

X 

xAir  Leaving  Fan  at  ty 
PL-  Pulley 

Foundation 
/     Floor  Line 

P 

^ 

Air  Filter 

~? 

I 

^ 

'2 

X 

Outside  Air      V 
Louvres  *    *~'\ 

Fan 

Outside  Wall-J^j 

By-pass  Damper 
FIG.  5.    COMBINATION  OF  RECIRCULATED  AIR  AND  OUTSIDE  AIR  (Case  C) 


Case  C.  (Fig.  5)  A  portion  of  the  air  circulated  is  recirculated  air  and  the  remainder, 
as  may  be  required  for  ventilating  purposes,  is  drawn  in  from  the  outside.  According  to 
Equations  4  and  5, 


The  temperature  of  the  resulting  mixture  of  outside  and  recirculated  air  entering  the 
tempering  coil  is: 

~»f     A  I  If    * 

(9) 


M 


Example  5.  Assuming  that  a  positive  supply  of  outside  air  (do  =  0.0864)  is  required 
for  ventilation  at  the  rate  of  90,000  cu  ft  per  hour  in  the  preceding  example,  then  M0 
-  0.0864  X  90,000  »  7776  Ib  per  hour  are  required,  measured  at  65  F. 

Mr  -  M  -  M0  -  38,889  -  7776  =  31,113  Ib 


Hi 


7776  X  0  +  31,113  X  65       K0  ^ 
k~  -  38^89  -  ~52F 

38,889  X  0.24  (140  -  52)  -  821,336  Btu. 


This  amount  of  work  may  be  accomplished  with  one  or  more  banks  of  heating  units, 
that  is,  either  a  single  reheater  or  a  tempering  coil  and  reheater. 


The  three  preceding  cases  refer  to  installations  in  which  conditioning 
the  air  to  maintain  certain  relative  humidity  requirements  does  not  enter 
into  the  problem,  as  for  example,  certain  types  of  industrial  installations. 
In  practically  all  modern  public  buildings,  theaters,  schools,  and  in  many 
industrial  installations,  the  ventilating  requirements  include  the  provision 
for  washing  and  humidifying  the  air  delivered  to  the  various  rooms  of  the 
structure. 

In  the  following  cases  it  is  assumed  that  in  addition  to  maintaining  a 
mean  room  temperature  t,  the  heating  and  ventilating  apparatus  is 
required  to  maintain  a  constant  relative  humidity  in  the  rooms. 

368 


CHAPTER  22 — FAN  SYSTEMS  OF  HEATING 


,/ Control  Valve 


Steam 


Rolling  Shutter- 
Outside  Air 
Louvres"^" 

Outside  Air  t 


1L 

•'       *      Steam          Control  Valve 
Tempering  Coil 

-  ... 

as: 

\                                 x 

jte        $^ 

1 

*w 

pashe|| 

"^Rehe 
sr 

->- 

^ 

ater 

f               *T^ 

1                    ^"^sprayWat, 

Fan 


4 


PURVIEW 


FIG.  6.   OUTSIDE  AIR  CIRCULATED;  CONSTANT  RELATIVE  HUMIDITY  IN  ROOM  (Case  D) 

Case  D.  (Fig.  6)  The  maximum  relative  humidity  that  may  be  maintained  within  the 
building  without  the  precipitation  of  moisture  on  single  glazed  sash  when  the  outside 
temperature  is  30  F  is  approximately  35  per  cent.  If  the  inside  temperature  t  is  70  F,  35 
per  cent  relative  humidity  corresponds  to  a  dew-point  temperature  of  41  F.  (See 
psychrometric  chart.) 

The  installation  shown  in  Fig.  6  contemplates  the  use  of  a  tempering  coil,  an  air 
washer  provided  with  a  water  heater,  and  a  reh eater.  The  tempering  coil,  one  section  in 
depth,  warms  the  incoming  air  to  approximately  35  F  to  prevent  freezing  any  of  the  spray 
water.  The  air  passing  through  the  spray  chamber  is  saturated  and  leaves  at  a  tempera- 
ture of  /i  =  41  F. 

The  heat  to  be  supplied  the  reheater  is: 

#1  =  0.24  (4  —  41)  M  Btu  per  hour. 

The  heat  to  be  supplied  the  tempering  coil  is: 

Hi  =  0.24  (35  -  t0)M  Btu  per  hour. 

The  amount  of  heat,  per  pound  of  air  circulated,  to  be  supplied  the  humidifying  washer 
or  humidifier  is  the  difference  between  the  heat  content  of  the  assumed  dry  air  entering 
the  washer  at  a  temperature  of  fw  =  35  F  and  the  leaving  saturated  air  at  t\  =  41  F 
(Chapter  1),  or: 

15.7  —  8.4  =  7.3  Btu  per  pound  of  dry  air. 
The  amount  of  heat  required  for  the  washer  is: 

Ha  =  7.3  M  Btu  per  hour. 
The  total  amount  of  heat  required  by  the  apparatus  is,  therefore: 

Hi  -f-  H3  +  H3  Btu  per  hour. 

If  a  washer  having  a  humidifying  efficiency  of  67  per  cent  without  water  heater  is  em- 
ployed it  will  be  necessary  to  heat  the  outside  air  drawn  into  the  apparatus  by  means  of 
the  tempering  and  preheater  coils  to  such  a  temperature  that  the  air  in  passing  through 

369 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

the  water  sprays  will  become  partially  saturated  (adiabatically)  having  a  moisture  con- 
tent per  pound  of  air  equal  to  saturated  air  at  41  F.  If  the  incoming  air  is  warmed  to 
£w  =  88  F  (requiring  a  two-section-depth  heating  unit)  it  will  be  cooled  in  the  washer  to 
64  F,  with  a  temperature  drop  of  88  -  64  =  24  deg. 

If  the  humidifying  efficiency  of  the  washer  were  100  per  cent,  the  air  would  become 
adiabatically  saturated  at  52  F  after  a  temperature  drop  of  88  —  52  =  36  F.  The 
efficiency  of  the  washer  is,  however,  only  67  per  cent,  so  that  the  actual  temperature  drop 
will  be  0.67  X  36  deg  or  24  deg,  as  used. 

The  heat  to  be  supplied  the  reheater  is  in  this  case  Hi  =  0.24  (k  -  64)  M  Btu  per 
hour,  and  the  heat  to  be  supplied  to  the  tempering  coil  and  preheater  is  H*  =  0.24 
(88  —  t0)  M.  The  total  heat  required  by  the  apparatus  is  Hi  +  H*,  no  heat  being 
supplied  to  the  washer. 


FIG.  7.     OUTSIDE  AIR  CIRCULATED;  CONSTANT  TEMPERATURE  AND  RELATIVE 
HUMIDITY  MAINTAINED  IN  EACH  ROOM  (Case  E) 


Case  E.     (Fig.  7)  The  temperature  ty  will  ordinarily  be  different  for  each  room 


With 


se  K.     (Fig.  7J    ine  temperature  ry  win 
H  and  M  fixed,  0.24  (ty  -  t}M  =  H,  or 


H 


0.24  M 


In  order  to  provide  the  proper  temperature  for  each  room,  a  booster  coil 
is  generally  installed  in  each  supply  duct  near  the  outlet  to  control  the  out- 
let temperature.^.  The  amount  of  steam  supplied  to  these  booster  units 
is  usually  controlled  automatically  by  individual  thermostats.  The  heat 
required  by  the  booster  coils  depends  on  the  temperature  range  through 
which  the  air  is  heated  and  the  quantity  of  air,  or 


0.24 


-  fe  -  tz}M 


(10) 


Total  Heat  to  be  Supplied 

The  total  heat  to  be  supplied  (JET)  is  equal  to  the  sum  of  the  heat 
requirements  of  the  various  heating  units  and  the  water  heater  of  the 
washer,  if  any,  plus  the  allowance  for  piping  tax.  (See  preceding  Cases 
A  to  E.) 


CHAPTER  22 — FAN  SYSTEMS  OF  HEATING 


Grate  Area,  Boiler  Selection 

The  required  grate  area  may  be  determined  by  the  following  formula: 


FXEXC 
where 

G  =  required  grate  area,  square  feet. 

F  —  calorific  value  of  fuel,  Btu  per  pound. 

C  —  combustion  rate,  pounds  per  square  foot  of  grate  per  hour. 

E  =  boiler  and  grate  efficiency,  per  cent. 

Example  6.  Using  the  data  in  Example  4,  and  assuming  coal  having  a  calorific  value 
of  12,000  Btu  per  pound,  a  combustion  rate  of  7  Ib  per  square  foot,  and  a  performance 
efficiency  of  0.60,  and  neglecting  the  piping  tax. 

r__     ^1,306,670        .         - 


~        12,000  X  0.60  X  7   ~  — H- 
Weight  of  Condensate 

The  normal  weight  of  condensate  to  be  handled  from  central  fan  sys- 
tems may  be  estimated  by  means  of  the  following  formula : 


'where 


_  60  X  Q  X  A* 
W  55.2  X  hfg 


W  =  weight  of  condensate,  pounds  per  hour. 
Q  —  total  volume  of  air,  cubic  feet  per  minute. 
AJ  =  temperature  rise  of  air,  degrees  Fahrenheit. 
Afg  —  latent  heat  of  steam  in  the  system,  Btu  per  pound. 

Ducts  and  Outlets,  Air  Filters,  Air  Washers 

The  design  of  the  duct  system  should  be  based  on  data  contained  in 
Chapter  20.  Air  washers  and  humidifiers  are  described  in  Chapter  11. 
For  information  on  air  filters,  see  Chapter  16. 

Static  Pressure 

The  total  static  pressure  against  which  the  system  must  operate  may 
be  found  by  summing  up  the  static  losses  through  the  complete  system 
from  the  outside  air  intake  to  the  discharge  outlets  or  nozzles.  This 
means  that  the  loss  due  to  friction  must  be  determined  for  each  piece  of 
apparatus  involved.  Most  of  these  values  may  be  obtained  from  manu- 
facturers' data  tables.  For  a  simple  system,  the  following  static  pressure 
drops  may  be  assumed  : 

1.  Outside  air  inlet,  comprised  of  screen,  louver  and  short  duct,  may  have  a  loss  of 
0.2  in.  of  water. 

2.  A  typical  oil  filter  at  rated  capacity  and  velocity  has  a  drop  of  0.25  in.  of  water. 

3.  The  loss  of  one  row  of  a  standard  make  tempering  stack  equals  0.09  in.  water. 

4.  The  loss  of  one  row  of  a  standard  make  preheater  equals  0.10  in.  water. 

5.  A  standard  humidifier  at  rated  velocity  may  have  a  loss  of  about  0,35  in.  water. 

6.  The  loss  through  one  row  of  a  standard  make  reheater  equals  0.12  in.  water. 

7.  A  fair  assumption  for  duct  losses  on  a  simple  system  is  0.25  in.  water. 

8.  The  static  pressure  for  a  nozzle  type  outlet  may  be  taken  as  0.1  in.  water. 

371 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

The  sum  of  these  values  equals  0.2  +  0.25  +  0.09  +  0.10  +  0.35 
+  0.12  +  0.25  +  0.1  =  1.46  in.  which  is  the  static  pressure  against  which 
the  system  must  operate. 

Fans  and  Control 

The  selection  of  fans  and  motors  may  be  based  on  data  contained  in 
Chapter  17.  Because  centrifugal  fans  reach  their  maximum  efficiency 
when  working  against  the  resistance  offered  by  the  average  central  fan 
heating  system,  they  are  well  adapted  to  such  systems  and  are  generally 
used.  Information  on  temperature  control  for  central  fan  systems  is 
given  in  Chapter  14. 

PROBLEMS  IN  PRACTICE 

1  •  What  are  the  functions  of  (a)   tempering  coils,   (b)  preheating  coils,   (c) 
reheating  coils,  (d)  hooster  coils,  (e)  water  heaters? 

a.  Tempering  coils  raise  the  temperature  of  incoming  air  above  the  freezing  point  of 
water. 

b.  Preheating  coils  add  to  the  air  sufficient  sensible  heat  above  the  dew  point  of  the 
conditioned  space  to  evaporate  the  amount  of  spray  water  required  for  humidification  . 
They  are  used  with  humidifying  type  air  washers. 

c.  Reheating  coils  raise  the  air  temperature  from  the  dew  point  to  approximately  the 
proper  delivery  temperature. 

d.  Booster  units  are  used  for  more  refined  individual  room  temperature  control. 

e.  Water  heaters  may  be  used  in  place  of  preheaters.    The  latent  heat  of  evaporation 
is  then  supplied  directly  to  the  water, 

2  •  What  saving  results  from  recirculating  some  of  the  room  air  and  reducing 
the  amount  of  outside  air? 

Because  outside  air  must  be  heated  to  room  temperature,  reducing  the  amount  of  outside 
air  produces  a  proportionate  saving  in  heat  or  fuel. 

3  •  What  items  make  up  the  total  heating  load  in  a  central  fan  heating  system? 

1.  The  net  heat  loss  from  the  conditioned  space. 

2.  The  heat  required  for  evaporation  of  water  for  humidification. 

3.  The  heat  required  to  raise  the  temperature  of  outside  air  to  room  temperature, 

4.  Heat  losses  from  pipes  and  ducts. 

4  •  Why  is  it  necessary  to  determine  the  total  static  pressure  of  a  central  fan 
heating  system? 

To  select  a  fan  of  maximum  efficiency  and  to  determine  the  power  required  to  operate 
the  fan. 

5  •  A  group  of  three  drafting  rooms,  having  a  total  volume  of  27,000  cu  ft,  a 
transmission  loss  of  110,100  Btuper  hour,  and  an  infiltration  loss  of  34,200  Btu  per 
hour  on  the  basis  of  0  F  outdoors  and  70  F  room  temperature,  is  to  be  heated  by 
a  recirculating  hot  blast  heating  system  with  air  entering  the  rooms  at  116  F. 
How  many  cubic  feet  per  minute,  measured  at  70  F,  will  be  required? 

Substitute  in  Equation  3.    H  =  110,100  +  34,200  =  144,300  Btu  per  hour;  ty  =  116  F; 
,  =  70  F;  Q  =  =  55l°°  -  173,160  cu  ft  per  hour. 


eta  =  2886. 


372 


CHAPTER  22 — FAN  SYSTEMS  OF  HEATING 


6  •  In  the  preceding  question,  if  the  hot  air  loses  4  F  between  heater  and 
rooms,  how  many  pounds  of  steam  per  hour  at  1-lh  gage  will  the  heating 
sections  condense? 

Substitute  in  Equation  12.  Q  =  2886  cfm,  from  solution  of  Question  5;  At  =  116  -f  4 
-  70  =  50  F;  hfg  =  968  Btu,  from  steam  table  in  Chapter  1. 

...       60  X  Q  X  At        60  X  2886  X  50       1tt0  ..          , 

W  "      55.2  X  frg      -      55.2  X  968       =  162  lb  per  hour' 

7  •  The  same  rooms  are  converted  to  chemical  laboratories,  requiring  the  intro- 
duction of  12  changes  of  outside  air,  measured  at  70  F,  per  hour  to  permit  the 
exhaust  fans  connected  to  the  chemical  hoods  to  maintain  only  a  slight  nega- 
tive pressure  in  the  rooms.    At  what  temperature  must  the  air  enter  the  rooms 
to  maintain  70  F  with  0  F  outside? 

Substitute  in  Equation  2.  H  =  110,100  +  34,200  =  144,300  Btu  per  hour;  Q  =  12  X 
27,000  -  324,000  Btu  per  hour;  *  =  70  F;  ty  -  ^?  +  1  =  55'2Q^^3°°  +  70 

(j/  O^4r,UUU 

=  94.6  F. 

8  •  In  the  preceding  question,  if  the  air  drops  2  F  between  the  heater  and  the 
rooms,  how  many  pounds  of  steam  per  hour  at  1-lb  gage  will   the  heating 
system  condense? 

Substitute  in  Equation  12.     Q  =  5400  cfm;  At  =  94.6  -f  2  =  96.6  F,  from  solution  of 

Question  7;  hfs  —  968  Btu,  from  steam  table  in  Chapter  1. 

w        60  X  Q  X  At         60  X  5400  X  96.6  ,          , 

W  =      55.2  X  feg      =         55.2  X  968         "  585  lb  per  hour. 

9  •  The  combination  hot  blast  heating  and  ventilating  system  for  the  dining 
rooms  of  a  hotel  is  to  heat  the  rooms  to  70  F  with  0  F  outside,  and  permit 
the  exhaust  fan  from  the  adjoining  kitchen  to  draw  5000  cfm  from  the  dining 
rooms.    The  transmission  losses  from  the  dining  rooms  total  240,000  Btu  per 
hour.    The  infiltration  into  the  dining  rooms  amounts  to  1000  cfm  from  out- 
doors and  1000  cfm  from  heater  rooms.     How  many  cubic  feet  per  minute, 
measured  at  70  F,  must  be  supplied  the  dining  rooms  if  the  air  enters  at  112  F? 

First  find  the  infiltration  loss  by  substituting  in  Equation  1. 

t  =  70  F;  to  =  0;  M0  =  d  X  Q  =  0.07495  X  60  X  1000  =  4497  lb  per  hour.  In  this  case 
d  and  Q  are  figured  at  70  F.  H0  ==  0.24  (t  -  /0)  ;  M0  =  0.24  (70  -  0)  X  4497  =  75,550 
Btu  per  hour. 

Next  by  substituting  in  Equation  3,  find  the  cubic  feet  per  hour  to  be  circulated.  H  = 
sum  of  transmission  and  infiltration  losses  in  room  =  240,000  -f  75,550  =  315,550  Btu 


per  hour;  fc  -  112  F;*  -  70F;£  -  -  =  55f  -  414,700  cu  ft  per  hour. 

ty  —  t  LL£  —  /u 

cfm  _  «g9?.  =  6912 

10  •  In  Question  9,  3000  cfm  of  outside  air  will  be  drawn  in  by  the  supply  fan 
and  3912  cfm  will  be  recirculated.  What  will  be  the  output  of  the  heating 
sections  in  Btu  per  hour  if  there  is  a  loss  of  2  F  between  the  heaters  and  the 
room? 

The  average  temperature  of  the  mixture  of  outdoor  and  recirculated  air  entering  the 

heater  -  30Q°  X  ^  ^^  X  7°  =  39.6  F.    Air  leaves  the  heater  at  112  +  2  =  114  F. 
691.2 

Referring  to  Equation  12,  W  X  kfg  =  total  heat  required  per  hour  =  -  =g-=  -  -  =  H. 


55.2 

I  cfm;  At  =  H4  -  39.6  -  74.4  F.    H  -  60X6912^74.4 
per  hour. 


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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

11  •  When  the  outdoor  wet-  and  dry-bulb  temperatures  are  0  F,  a  certain  print- 
ing shop  is  to  be  maintained  at  75  F  and  40  per  cent  relative  humidity  by  means 
of  an  air  conditioning  system  having  tempering  sections,  an  air  washer,  and 
reheating  sections.     The  transmission  loss  is  80,000  Btu  per  hour  and  the 
infiltration  is  10,000  cu  ft  per  hour,  measured  at  0  F.    No  outside  air  connection 
is  provided.     How  many  pounds  of  air  per  hour  at  120  F  must  be  discharged 
to  the  shop? 

Infiltration  heat  loss,  by  Equation  1  =  H0  =  0.24  (t  -  t0]  MQ.  By  Equation  6,  M0  = 
d0Q0  =  0.08636  (from  Table  5,  Chapter  1)  X  10,000  =  863.6  Ib  per  hour;  t  =  75  F; 
to  =  0  F;  Ho  =  0.24  (75  -  0)  863.6  =  15,544  Btu  per  hour.  Total  heat  loss  in  room 
=  80,000  4-  15,544  =  95,544  Btu  per  hour  =  H. 

To  secure  the  total  weight  of  air  to  be  introduced  into  the  space,  substitute  in  Equation 

A    *,  H  95,544  00_  .,          , 

4'  M  =  0.24  (fr  -1)  ~  0.24(120-75)  =  8846  lb  pef  h°Ur' 

12  •  In  the  preceding  example:     (a)  How  many  Btu  per  hour  are  used  to  heat 
the  room?     (b)  How  many  pounds  of  water  must  be  evaporated  per  hour  to 
humidify  the  space?     (c)  How  many  Btu  will  be  required   to  evaporate   this 
water,  basing  the  latent  heat  of  evaporation  on  the  approximate  figure  of 
1050  Btu? 

a.  Btu  to  heat  room  ==  95,544  as  derived  in  preceding  solution. 

b.  Saturated  air  at  75  F  contains  0.01877  lb  of  water  vapor  per  pound  of  dry  air.     At 
40  per  cent  relative  humidity  the  air  would  contain  0.40  X  0.01877  =  0.00750  lb  of 
water  vapor  per  pound  of  dry  air;  at  0  F,  saturated  air  contains  0.00078  lb  of  water 
vapor  per  lb  of  dry  air.     The  amount  of  water  vapor  required  to  humidify  the  air  = 
0.00750  -  0.00078  =  0.00672  lb  per  cu  ft.    Infiltration  amounts  to  863.6  lb  per  hour  as 
derived  in  the  preceding  solution,  so  863.6  X  0.00672  =  5.80  lb  of  water  vapor  per  hour 
required. 

c.  The  heat  required  to  evaporate  this  water  =  5.80  X  1050  =  6090  Btu  per  hour. 


374 


Chapter  23 

MECHANICAL  WARM  AIR  FURNACE 

SYSTEMS 

Fan  Furnaces,  Fans  and  Motors,  Elimination  of  Noise,  Air  Washers 
and  Filters,  Cooling^  Methods,  Duct  Design,   Controls,  Selecting 
the  Furnace,  Selecting  the  Fan,  Humidity  Provision  for  Cooling      * 
System,  Heavy  Duty  Fan  Furnaces 

MECHANICAL  warm  air  or  fan  furnace  heating  systems,  which  are  a 
special  type  of  central  fan  systems,  are  particularly  adapted  to 
residences,  small  office  buildings,  stores,  banks,  schools,  and  churches. 
Circulation  of  air  is  effected  by  motor-driven  fans  instead  of  by  the 
difference  in  weight  between  the  heated  air  leaving  the  top  of  the  casing 
and  the  cooled  air  entering  its  bottom,  as  in  gravity  systems  described  in 
Chapter  24.  The  advantages  of  mechanical  systems,  as  compared  with 
gravity  systems  are: 

1.  The  furnace  can  be  installed  in  a  corner  of  the  basement,  leaving  more  basement 
room  available  for  other  purposes. 

2.  Basement  distribution  piping  can  be  made  smaller  and  can  be  so  installed  as  to 
give  full  head  room  in  all  parts  of  the  average  basement,  or  be  completely  concealed 
from  view  except  in  the  furnace  room. 

3.  Circulation  of  air  is  positive,  and  in  a  properly  designed  system  can  be  balanced  in 
such  a  way  as  to  give  a  greater  uniformity  of  temperature  distribution. 

4.  Humidity  control  is  more  readily  attained. 

5.  The  air  may  be  cleaned  by  air  washers  or  filters,  or  both. 

6.  Some  cooling  effect  in  summer  will  result  from   the   installation  of  a  properly 
designed  system- 

7.  The  fan  and  duct  equipment  may  be  utilized  for  a  complete  cooling  and  dehumidi- 
fying  system  for  summer,  using  either  ice,  mechanical  refrigeration,  or  low  temperature 
water  for  cooling  and  dehumidifying,  or  adsorbers  for  dehumidifying. 

8.  The  use  of  the  fan  increases  the  volume  of  air  which  can  be  handled,  thereby 
increasing  the  rate  of  heat  extraction  from  a  given  amount  of  heating  surface  and 
insuring  sufficient  air  volume  to  obtain  proper  distribution  in  a  large  room. 

Much  of  the  equipment  used  in  central  fan  systems  is  the  subject  matter 
of  other  chapters.  It  is  the  purpose  of  this  chapter  to  discuss  the  co- 
ordinated design  and  to  deal  in  detail  only  with  problems  not  covered 
elsewhere  which  refer  particularly  to  the  whole  problem  of  fan  warm  air 
furnace  heating  and  air  conditioning. 

FAN  FURNACES 

Furnaces  for  mechanical  warm  air  systems  may  be  made  of  cast-iron, 
steel,  or  alloy.  Cast-iron  furnaces  are  usually  made  in  sections  and  must 
be  assembled  and  cemented  or  bolted  together  on  the  job.  Steel  furnaces 
are  made  with  welded  or  riveted  seams.  The  proper  design  of  the  furnace 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

depends  largely  on  the  kind  of  fuel  to  be  burned.  Accordingly,  various 
manufacturers  are  making  special  units  for  coal,  oil  and  gas.  Each  type 
of  fuel  requires  a  distinct  type  of  furnace  for  highest  efficiency  and  econ- 
omy, substantially  as  follows: 

1.  Coal  Burning: 

a.  Bituminous — Large  combustion  space  with  easily  accessible  secondary  radiator 
or  flue  travel. 

b.  Anthracite  or  coke — Large  fire  box  capacity  and   liberal  secondary  heating 
surfaces. 

2.  Oil  Burning: 

a.  Liberal  combustion  space. 

b.  Long  fire  travel  and  extensive  heating  surface. 

3.  Gas  Burning: 

a.  Extensive  heating  surface. 

b.  Close  contact  between  flame  and  heating  surface. 

A  combustion  rate  of  from  5  to  8  Ib  of  coal  per  square  foot  of  grate  per 
hour  is  recommended  for  residential  heaters.    A  higher  combustion  rate  is 


FIG.  1.    USUAL  METHOD  OF  BAFFLING  ROUND  CASINGS  FOR  FAN  FURNACE  WORK 

A.    Liner,  1  in.  from  casing.      B.    Hole  to  vent  baffle. 
C.    Baffle,  closed  top  and  bottom.      D.    Outer  casing. 

permissible  with  larger  furnaces  for  buildings  other  than  residences, 
depending  upon  the  ratio  of  grate  surface  to  heating  surface,  firing  period, 
and  available  draft. 

Where  oil  fuel  is  used,  care  must  be  exercised  in  selecting  the  proper  size 
and  type  of  burner  for  the  particular  size  and  type  of  furnace  used.  It  is 
recommended  that  the  system  be  designed  for  blow-through  installations, 
so  that  the  furnace  shall  be  under  external  pressure  in  order  to  minimize 
the  possibility  of  leakage  of  the  products  of  combustion  into  tlie  air 
circulating  system. 

In  residential  furnaces  for  coal  burning,  the  ratio  of  heating  surface  to 
grate  area  will  average  about  20  to  1 ;  in  commercial  sizes  it  may  run  as 
high  as  50  to  1,  depending  on  fuel  and  draft.  Furnaces  may  be  installed 
singly,  each  furnace  with  its  own  fan,  or  in  batteries  of  any  number  of 
furnaces,  using  one  or  more  fans. 

Casings  are  usually  constructed  of  galvanized  iron,  26-gage  or  heavier, 
but  they  may  also  be  constructed  of  brick.  Galvanized  iron  casings  should 

376 


CHAPTER  23  —  MECHANICAL  WARM  AIR  FURNACE  SYSTEMS 


be  lined  with  black  iron  liners,  extending  from  the  grate  level  to  the  top  of 
the  furnace  and  spaced  from  1  in.  to  1  J^  in.  from  the  outer  casing.  Casings 
for  commercial  or  heavy  duty  furnaces,  if  built  of  galvanized  iron,  should 
be  insulated  with  fireproof  insulating  material  at  least  2-in.  thick.  It  is 
generally  believed  that  either  brick  or  sheet  metal  casing  should  be 
equipped  with  baffles  to  secure  impingement  of  the  air  to  be  heated 
against  the  heating  surfaces.  Brick  furnace  casings  should  be  supplied 
with  access  doors  for  inspection. 

For  furnace  casings  sized  for  gravity  flow  of  air,  where  a  fan  is  to  be 
used,  many  manufacturers  recommend  the  use  of  special  baffles  to  restrict 
the  free  area  within  the  casing  and  to  force  impingement  of  the  air  against 
the  heating  surfaces.  The  method  of  making  these  baffles  for  furnaces 
with  ^  top  horse-shoe  radiators  and  for  furnaces  with  back  crescent  radia- 
tors is  illustrated  in  Fig.  1. 

Either  square  or  round  casings  may  be  used.  Where  square  casings  are 
used,  the  corners  must  be  baffled  to  reduce  the  net  free  area  and  to  force 
impingement  of  air  against  the  heating  surfaces.  Fig.  2  shows  the  usual 
method  of  baffling  square  furnace  casings  for  fan-furnace  work. 


FIG.  2.  METHOD  OF  BAFFLING  SQUARE  FURNACE  CASING  FOR  FAN  FURNACE  WORK 


A .  Baffle,  closed  top  and  bottom, 
casing.       C.    Outer  casing.      D. 


B.  Liner,  1  in.  from 
Hole  to  vent  baffle. 


The  hood  or  bonnet  of  the  casing  above  the  furnace  should  be  as  high 
as  basement  conditions  will  allow,  to  form  a  plenum  chamber  over  the  top 
of  the  furnace.  This  tends  to  equalize  the  pressure  and  temperature  of  the 
air  leaving  the  bonnet  through  the  various  openings.  It  is  generally  con- 
sidered advisable  to  take  off  the  warm  air  pipes  from  the  side  of  the  bonnet 
near  the  top,  as  this  method  of  take-off  allows  the  use  of  a  higher  bonnet 
and  thus  provides  a  larger  plenum  chamber.  Fig.  3  illustrates  a  complete 
residence  fan  furnace  installation  showing  location  of  fan,  furnace,  filters, 
plenum  chamber  and  method  of  take-off  of  warm  air  pipe. 

FANS  AND  MOTORS 

Centrifugal  type  fans  are  most  commonly  used,  and  these  may  be 
equipped  with  either  backward  or  forward  curved  blades.  Low  tip  speed 
is  desirable  for  the  elimination  of  air  noise,  especially  where  forward 
curved  blades  are  used.  Motors  may  be  mounted  on  the  fan  shaft  or 
outside  of  the  fan  with  belt  connection.  Multi-speed  motors  or  pulleys 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

are  desirable  to  provide  a  factor  of  safety  and  to  allow  for  more  rapid 
circulation  for  summer  cooling. 

For  additional  information  on  fans  and  motors,  see  Chapter  17. 

NOISE  ELIMINATION 

Special  attention  must  be  given  to  the  problem  of  noise  elimination. 
The  fan  housing  must  not  be  directly  connected  with  metal,  either  to  the 
furnace  casing  or  to  the  return  air  piping.  It  is  common  practice  to  use 
canvas  strips  in  making  these  connections.  Motors  and  their  mountings 


FIG.  3.   COMPLETE  RESIDENCE  FAN  FURNACE  INSTALLATION  SHOWING  LOCATION  OF  FAN, 
FURNACE,  FILTERS,  PLENUM  CHAMBER  AND  METHOD  OF  TAKE-OFF  OF  WARM  AIR  PIPE 


A.  Transition  fitting. 

B.  Filters. 

C.  Capped  opening. 

D.  Canvas  connection. 

E.  Pulley — 3  diam.  V-type. 


F.  Eliminator. 

G.  Solenoid  valve. 
H.  Pressure  gage. 
J.   Water  supply. 
K.  Drain. 


must  be  carefully  chosen  for  quiet  operation.  Electrical  conduit  and 
water  piping  must  not  be  fastened  to,  nor  make  contact  with,  fan  housing. 
The  installation  of  a  fan  directly  under  a  cold  air  grille  is  not  recommended 
on  account  of  the  noise  objection.  See  also  Chapter  18. 

AIR  WASHERS  AND  FILTERS 

Washers  for  residence  systems  may  be  provided  in  separate  housings 
to  be  installed  on  the  inlet  or  outlet  side  of  the  fan,  or  they  may  be 
integral  with  the  fan  construction.  They  operate  at  water  pressures  of 
from  10  to  30  Ib  and  use  two  or  more  spray  nozzles  for  washing  and 
humidification.  The  sprays  should  be  adjusted  to  completely  cover  the 
air  passages. 

Washers  are  usually  controlled  by  solenoid  valves  wired  in  parallel  with 

378 


CHAPTER  23 — MECHANICAL  WARM  AIR  FURNACE  SYSTEMS 

the  fan  motor.  The  water  supply  may,  in  turn,  be  controlled  by  a 
humidity-controlling  device  located  in  one  of  the  living  rooms,  so  that  the 
washer  will  operate  at  all  times  when  the  fan  is  in  operation,  unless  the 
relative  humidity  should  rise  beyond  a  desirable  percentage.  Washers 
used  in  connection  with  commercial  or  heavy  duty  plants  should  be  a 
regulation  type  of  commercial  washer. 

There  are  many  satisfactory  types  of  filters  on  the  market.  These 
include  dry  filters,  viscous  filters,  oil  filters  and  other  types,  some  of  which 
must  be  cleaned,  some  of  which  must  be  cleaned  and  recharged  with  oil, 
and  some  of  which  are  inexpensive  and  may  be  discarded  when  they 
become  dirty,  and  replaced  with  new  ones. 

The  resistance  of  a  filter  must  be  considered  in  the  design  of  the  system 
since  the  resistance  rises  rapidly  as  the  filter  becomes  dirty,  thus  im- 
pairing the  heating  efficiency  of  the  furnace,  in  fact,  endangering  the  life 
of  the  furnace  itself.  Manufacturers'  ratings  of  filters  must  be  carefully 
regarded,  and  ample  filter  area  must  be  provided.  Filters  must  be 
replaced  or  cleaned  when  dirty.  See  also  Chapter  16. 

COOLING  METHODS 

Some  cooling  may  be  obtained  under  certain  conditions  by  the  use  of 
basement  air.  A  more  positive  cooling  effect  may  be  obtained  through  air 
washers  where  the  temperature  of  the  water  is  sufficiently  low  (55  F  or 
lower),  and  where  a  sufficient  volume  of  water  can  be  provided.  Unless 
the  water  is  below  the  dew  point  temperature  of  the  indoor  air  at  the  time 
the  washer  is  started,  both  the  relative  and  absolute  humidities  will  be 
somewhat  increased. 

Coils  of  copper  finned  tubing  through  which  cold  water  is  pumped  are 
available  for  cooling.  They  require  less  space  than  air  washers  and  have 
the  advantage  that  no  moisture  is  added  to  'the  air  when  the  temperature 
of  the  water  rises  above  the  dew  point.  Ample  coil  surface  is  necessary 
with  this  type  of  cooling. 

It  is  thoroughly  feasible  to  use  ice  or  mechanical  refrigeration  in  con- 
nection with  the  fan  and  duct  system  for  the  heating  installation,  and  to 
cool  the  building  by  this  method,  provided  the  building  is  reasonably 
well  constructed  and  insulated.  Windows  and  doors  should  be  tight,  and 
awnings  should  be  supplied  on  the  sunny  side  of  the  building.  See  also 
Chapters  9  and  10. 

Study  of  these  problems  sponsored  by  the  AMERICAN  SOCIETY  OF 
HEATING  AND  VENTILATING  ENGINEERS  in  cooperation  with  the  National 
Warm  Air  Heating  Association  is  in  progress  at  the  University  of  Illinois. 
The  following  conclusions  may  be  drawn  from  the  studies  thus  far  com- 
pleted, subject  to  the  limitations  of  the  conditions  under  which  the  tests 
were  run1: 

1.  An  uninsulated  building  of  ordinary  residential  type  may  require  the  equivalent  of 
three  tons  of  ice  in  24  hours  on  days  when  the  maximum  outdoor  temperature  reaches 
100  F  if  an  effective  temperature  of  approximately  72  deg  is  maintained  indoors. 


*See  A.S.H.V.E.  research  paper  entitled  Study  of  Summer  Cooling  in  the  Research  Residence  at  the 
University  of  Illinois,  by  A.  P.  Kratz  and  S.  Konzo  (A.  S.  H,  V.  E.  Journal  Section  Heating,  Piling  an$  Air 
Conditioning,  February,  1933). 

379 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

2.  The  use  of  awnings  at  all  windows  in  east,  south,  and  west  exposures  may  result  in 
savings  of  from  20  to  30  per  cent  in  the  required  cooling  load. 

3.  The  cooling  load  per  degree  difference  in  temperature  is  not  constant  but  increases 
as  the  outdoor  temperature  increases. 

4.  The  heat  lag  of  the  building  complicates  the  estimation  of  the  cooling  load  under 
any  specified  conditions  and  makes  such  estimates,  based  on  the  usual  methods  of  com- 
putation, of  doubtful  value. 

5.  The  seasonal  cooling  requirements  are  extremely  variable  from  year  to  year,  and 
the  ratio  between  the  degree-hours  of  any  two  seasons  occurring  within  a  10-year  period 
may  be  as  high  as  7.5  to  1.    Hence  an  average  value  of  the  degree-hours  cooling  per 
season  is  comparatively  meaningless. 

6.  The  results  of  the  tests  suggest  the  use  of  a  fan  at  night  either  to  provide  more 
comfortable  conditions  during  the  following  day  without  provision  for  cooling,  or  to 
reduce  the  load  required  for  cooling  during  the  following  day.    Experience  has  shown  that 
the  volume  of  air  required  for  cooling,  depending  upon  the  climate  and  the  construction 
of  the  building,  must  usually  be  from  50  to  100  per  cent  greater  than^that  required  for 
heating.    If  the  size  of  the  fan  is  based  upon  the  summer  requirement,  its  output  may  be 
reduced  sufficiently  to  meet  winter  heating  needs. 

7.  Attic  exhaust  fans  are  becoming  popular  adjuncts  for  night  duty.  (See  Chapter  13.) 

DUCT  DESIGN 

The  ducts  may  be  either  round  or  rectangular.  Rectangular  ducts 
should  be  as  nearly  square  as  possible;  the  width  should  not  be  greater 
than  four  times  the  breadth.  The  radii  of  elbows  should  be  not  less  than 


]/OL  UMUL 


FIG.  4.   THREE  TYPES  OF  DAMPERS  COMMONLY  USED  FOR  TRUNK  AND  INDIVIDUAL 

DUCT   SYSTEMS 

one  and  one-half  times  the  pipe  diameter  for  round  pipes,  or  the  equiva- 
lent round  pipe  size  in  the  case  of  rectangular  ducts. 

The  ducts  or  piping  may  be  designed  either  as  a  trunk  line  system  or  as 
a  system  of  individual  ducts  from  the  furnace  casing  to  each  register.  The 
engineering  problems  incident  to  the  design  of  a  trunk  line  system  are 
somewhat  more  difficult  than  for  the  individual  duct  system.  The  trunk 
line  system  is  generally  a  tailor-made  job,  whereas  the  individual  duct 
system  with  which  either  round  or  square  ducts  may  be  used  may  fre- 
quently be  assembled  from  stock  materials  and  thus  installed  at  a  con- 
siderable saving.  Individual  ducts  may  frequently  be  grouped  to  simulate 

380 


CHAPTER  23 — MECHANICAL  WARM  AIR  FURNACE  SYSTEMS 


a  trunk  duct  system  in  appearance.    The  design  of  ducts  for  air  flow  is 
described  in  Chapter  20. 

Dampers 

Suitable  dampers  are  essential  to  any  trunk  or  Individual  duct  system, 
as  it  is  virtually  impossible  to  so  lay  out  a  system  that  it  will  be  absolutely 
in  balance  without  the  use  of  dampers.  Special  care  must  be  used  in  the 
design  of  any  system  to  avoid  turbulence  and  to  minimize  resistance. 
Sharp  elbows,  angles,  and  offsets  should  be  avoided.  See  Figs.  1  and  2f 
Chapter  20. 

Three  types  of  dampers  are  commonly  used  in  trunk  and  individual 
duct  systems.  Volume  dampers  are  used  to  completely  cut  off  or  reduce 
the  flow  through  pipes.  (See  A  and  2?,  Fig.  4.)  Splitter  dampers  are  used 
where  a  branch  is  taken  off  from  a  main  trunk.  (See  C,  Fig.  4.)  Squeeze 
dampers  are  used  for  adjusting  the  volume  of  air  flow  and  resistance 
through  a  given  duct.  (See  D,  Fig.  4.)  It  is  essential  that  a  damper  be 
provided  for  each  main  or  duct  branch.  A  positive  locking  device  should 
be  used  with  each  type  of  damper. 

Supply  and  Return  Air  Registers 

Supply  registers  located  in  the  floor  are  effective,  but  as  they  require 
frequent  attention  to  keep  them  clean  they  should  be  avoided  where 
another  effective  register  location  can  be  found.  Unless  registers  located 
in  the  baseboard  are  well  proportioned  and  designed  to  harmonize  with 
the  trim,  they  may  be  unsightly.  Registers  which  are  located  in  side  walls 
above  the  baseboard  or  in  the  ceiling  should  be  of  an  effective  air-diffusing 
type.  All  registers  should  be  sealed  against  leakage  around  the  borders 
or  margins. 

Velocities  through  registers  may  be  reduced  by  the  use  of  registers 


FIG.  5. 


DIFFUSERS  IN  TRANSITION  FITTINGS  TO  EQUALIZE  VELOCITIES 
THROUGH  REGISTER  FACES 


larger  than  the  connecting  pipes.  Some  suggestions  for  equalizing  veloci- 
ties over  the  face  area  of  the  register  by  means  of  diffusers  are  illustrated 
in  Fig.  5.  Merely  to  use  a  larger  register  may  not  result  in  materially 
reduced  velocities  unless  such  diffusers  are  used. 

Care  should  be  exercised  in  making  the  connection  between  the  supply 
register  and  its  box  to  prevent  streaking  of  the  wall.  All  warm  air 
registers  should  be  equipped  with  dampers  or,  better,  with  diffuser 
dampers  which  may  be  used  to  direct  air  currents  in  such  a  way  that  they 
will  not  be  objectionable.  (See  Chapter  19.) 

381 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


ko  £ 


382 


CHAPTER  2£— MECHANICAL  WARM  AIR  FURNACE  SYSTEMS 


CONTROLS 

Air  stratification,  high  bonnet  temperatures,  excessive  flue  gas  tem- 
peratures, and  heat  overrun  or  lag  in  the  system  can  be  largely  elimi- 
nated through  proper  care  in  the  planning  and  installation  of  the  control 
system.  The  essential  requirements  of  the  control  are: 

1.  To  keep  the  fire  burning  when  using  solid  fuel  regardless  of  the  weather. 

2.  To  avoid  excessive  bonnet  temperatures  with  resultant  radiant  heat  losses  into  the 
basement. 

3.  To  avoid  the  overheating  of  certain  rooms  through  gravity  action  during  off 
periods  of  blower  operation. 

4.  To  have  a  sufficient  supply  of  heat  available  at  all  times  to  avoid  lag  when  the 
room  thermostat  calls  for  heat. 

5.  To  prevent  cold  air  delivery  when  heat  supply  is  insufficient. 

6.  To  avoid  heat  loss  through  the  chimney  by  keeping  stack  temperatures  low. 

7.  To  provide  quick  response  to  the  thermostat,  with  protection  against  overrun. 

8.  To  provide  for  humidity  control. 

9.  To  provide  a  means  of  summer  control  of  cooling. 
10.  To  protect  against  fire  hazards. 

The  following  controls  are  desirable: 

1.  A  thermostat  located  at  a  point  where  maximum  fluctuation  in  temperature  can  be 
expected,  in  order  to  secure  frequent  operation  of  fans,  drafts,  and  burners.    This  location 
would  be  near  an  outside  wall  but  not  upon  it7  in  a  sun  room,  or  in  a  room  with  some 
unusual  exposure.    The  thermostat,  of  course,  should  not  be  located  where  it  will  be 
affected  by  direct  radiant  heat  from  the  sun  or  from  a  fireplace,  or  by  direct  heat  from 
any  warm  air  duct  or  register. 

2.  A  furnacestat  located  in  the  bonnet  to  permit  blower  operation  only  between  the 
temperatures  of  100  F  and  150  F.    In  certain  extreme  cases  it  may  be  necessary,  or 
weather  conditions  may  make  it  advisable,  to  adjust  the  high  limit  to  a  higher  tempera- 
ture than  that  given.    Another  location  sometimes  used  for  the  furnacestat  is  in  the  main 
duct  near  the  frame  opening  from  the  bonnet. 

3.  A  protective  limit  control  located  in  the  bonnet  to  shut  down  the  system  inde- 
pendently of  the  thermostat  if  the  bonnet  temperature  exceeds  225  F. 

4.  On  oil  and  §as  burner  installations,  a  control  is  usually  included  which  will  shut 
down  the  system  if  the  fire  goes  out  or  if  there  is  a  failure  of  the  ignition  system. 

5.  A  humidistat  to  regulate  the  moisture  supplied  to  the  rooms. 

6.  On  automatic  stoker  installations,  a  control  is  usually  included  which  will  start 
the  operation  regardless  of  thermostat  settings  whenever  the  bonnet  temperature 
indicates  that  the  fire  is  dying. 

While  it  is  usually  all  right  to  start  and  stop  the  fan  in  residential  con- 
trol work  and  in  auditoriums  and  other  places  where  many  people  may 
gather,  the  fan  should  as  a  rule  be  allowed  to  run  continuously  and  the 
control  should  be  cared  for  in  other  ways. 

SELECTING  THE  FURNACE 

The  following  formula  may  be  used  to  compute  the  grate  area  of  a 
residence  "furnace,  assuming  a  ratio  of  heating  surface  to  grate  area 
of  20  to  1 : 

7-7-. 

G  =   FXCXE  (1) 

383 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


o  o 
IB 


SSRHg 

gj    &•    tH    M  ~ 

sSESp 

^Se5 

w  B  g  ««  w  w 

w  J3  S  iJ  S  a: 
£5^  O^  S  g 

8O.  W       *4  ffl 
nS'S^g 

ili<§5 


384 


CHAPTER  23 — MECHANICAL  WARM  AIR  FURNACE  SYSTEMS 


where 


G  =  required  grate  area,  square  feet. 

H  =  total  heat  loss  from  building,  Btu  per  hour. 

F  =  calorific  value  of  coal,  Btu  per  pound. 

C  —  combustion  rate  in  pounds  of  fuel  per  square  foot  of  grate  per  hour. 

E  —  furnace  efficiency  based  on  heat  available  at  register  faces. 

In  practice  it  is  customary  to  use  the  following  constants : 

F  =  13,000  (For  specific  values,  see  Table  1,  Chapter  27). 
C  =  5  to  10  Ib  (Use  8  Ib  as  maximum  in  residence  work). 

E  =*  55  per  cent  to  65  per  cent  depending  on  fuel  burned.    Lower  efficiency  must 
be  used  with  highly  volatile  solid  fuel. 

Where  ratio  of  heating  surface  to  grate  area  is  less  or  greater  than  20  to 
1,  deduct  or  add  2  per  cent  from  or  to  rating  of  furnace  for  each  unit 
decrease  or  increase  in  ratio,  as  the  case  may  be.  The  foregoing  procedure 
for  determining  the  size  of  the  furnace  to  be  used  applies  to  continuously 
heated  buildings. 

Although  intermittently  heated  buildings  usually  have  their  heat  losses 
computed  according  to  the  standard  rules  for  determining  such  losses, 
these  rules  do  not  take  into  account  the  heat  which  will  be  absorbed  by 
the  cold  material  of  the  building  after  the  air  is  raised  in  temperature. 
This  heat  absorption  must  be  added  to  the  normal  heat  loss  of  the  building 
to  determine  the  load  which  the  heating  plant  must  carry  through  the 
warming-up  process.  It  is  customary  to  increase  the  normal  heat  loss 
figure  by  from  50  to  150  per  cent  depending  upon  the  heat  capacity  of  the 
construction  material,  the  higher  percentage  applying  to  materials  of 
high  heat  capacity  such  as  concrete  and  brick.  Fan  furnace  systems  are 
well  adapted  for  heating  intermittently  heated  buildings  as  these  systems 
do  not  require  the  warming  of  intermediate  piping,  radiators,  or  con- 
vectors,  the  generation  of  steam,  or  the  heating  of  hot  water. 

Follow  the  same  methods  for  an  oil  furnace  as  for  coal  where  a  con- 
version unit  is  to  be  used,  making  sure  that  the  ratio  of  heating  surface  to 
grate  area  exceeds  20  to  1.  If  it  does  not,  a  size  larger  furnace  should  be 
selected.  Use  the  manufacturers'  Btu  ratings  of  furnaces  designed  for 
exclusive  use  with  oil,  and  select  a  burner  with  liberal  excess  capacity. 

The  selection  of  the  proper  size  gas  furnace  for  a  constantly  heated 
building  can  be  easily  made  by  using  the  following  American  Gas  Associa- 
tion formula : 

*  =  <§  (2) 

where 

H  =  total  heat  loss  from  building  in  Btu  per  hour. 

R  =  official  A.G.A.  output  rating  of  the  furnace  in  Btu  per  hour. 

In  the  case  of  converted  warm  air  furnaces  a  slightly  different  procedure 
is  necessary,  as  the  Btu  input  to  the  conversion  burner  must  be  selected 
rather  than  the  furnace  output.  The  proper  sizing  may  be  done  by  means 
of  the  following  formula: 

/  =  1.56H  (3) 

385 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 
where    ' 

I  =  Btu  per  hour  input. 

The  factor  1.56  is  the  multiplier  necessary  to  care  for  a  10  per  cent  heat 
loss  in  the  distributing  ducts  and  an  efficiency  of  70  per  cent  in  the  con- 
version burner. 

SELECTING  THE  FAN 

Choose  a  fan  which,  according  to  its  manufacturer's  rating,  is  capable 
of  delivering  a  volume  of  air,  expressed  in  cubic  feet  per  minute,  against 
a  frictional  resistance,  expressed  in  inches  of  water,  computed  by  adding 
together  the  following  items: 

1.  The  frictional  resistance  of  a  warm  air  trunk  or  leader. 

2.  The  frictional  resistance  of  a  return  air  trunk  or  duct. 

3.  The  resistance  to  the  flow  of  total  volume  of  air  through  the  furnace  casing  or  hood, 
which  is  usually  considered  from  0.10  to  0.15  inches  of  water. 

4.  The  frictional  resistance  through  any  other  accessories,  such  as  washers  or  niters. 

5.  A  factor  of  safety  of  10  per  cent  of  the  resistance  calculated  above. 

HUMIDITY 

Mechanical  warm  air  systems  offer  an  excellent  means  of  proportioning 
and  distributing  moisture-bearing  air;  consequently,  during  the  winter 
months  humidifiers  may  be  employed  to  deliver  water  vapor  to  the  fan- 
driven  air  stream  in  proper  amounts  to  produce  a  more  humid  atmos- 
phere, with  increased  comfort  for  people  and  increased  life  for  household 
furnishings.  Temperatures  and  relative  humidities  should  be  governed 
within  the  limits  of  the  generally  accepted  standards.  See  Chapters  2  and 
3  for  more  detailed  information  on  this  point. 

In  earlier  types  of  furnaces,  water  evaporating  pans  were  usually  placed 
in  the  cool  portions  of  the  air  stream,  but  modern  types  usually  locate 
them  in  air  which  has  been  heated  by  contact  with  the  heating  surfaces. 
To  change  water  into  vapor  capable  of  being  carried  in  an  air  stream  as 
part  of  the -mixture,  about  1000  Btu  per  pound  are  required.  Without 
the  addition  of  this  heat,  termed  the  latent  heat  of  evaporation,  water, 
injected  intathe  air  will  be  carried  along  in  the  form  of  tiny  globules  until 
it  falls  out  of  the  stream  or  is  deposited  upon  some  surface.  . 

Furthermore,  when  dry  air  is  in  contact  with  water  for  a  sufficient 
length  of  time  without  the  presence  of  a  sizable  body  of  water  or  a  source 
other  than  air  from  which  this  latent  heat  of  evaporation  can  be  taken, 
such  heat  is  supplied  from  the  air.  There  is,  therefore,  a  trend  in  present 
practice  toward  heating  the  water  in  addition  to  heating  the  air.  Equip- 
ment for  doing  this  may  make  use  of  sprays,  or  it  may  take  the  form  of 
water  circulating  coils  placed  within  the  combustion  chamber  and  con- 
nected by  pipes  to  the  humidifier  pans  where  a  constant  water  level  is 
maintained  by  some  separate  float  device.  (See  Chapter  11.) 

PROVISION  FOR  COOLING  SYSTEM 

If  the  system  is  to  be  used  for  cooling,  the  following  provisions  should 
be  made: 


CHAPTER  23 — MECHANICAL  WARM  AIR  FURNACE  SYSTEMS  ' 

1.  Where  cooling  is  to  be  secured  through  air  circulation  only: 

a.  Provide  for  an  increase  of  50  to  100  per  cent  in  fan  capacity  through  multi-speed 
pulleys  or  other  means, 

b.  If  basement  air  or  outside  night  air  is  to  be  used,  provide  suitable  basement 
opening  in  duct  system,  or  outdoor  air  intake, 

2.  Where  water  below  55  F  or  artificial  refrigeration  or  ice  is  to  be  used: 

a.  Provide  outside  air  duct  for  circulation  of  cool  night  air  for  economy. 

b.  Make  provision  in  return  duct  system  for  cooling  unit. 

c.  Make  provision  for  control  of  the  fan  speed,  during  winter  operation,  to  give 
a  sufficient  and  draftless  air  movement. 

HEAVY  DUTY  FAN  FURNACES 

Fan  furnaces  for  large  commercial  and  industrial  buildings  are  available 
in  sizes  ranging  from  400,000  to  3,000,000  Btu  per  hour  per  unit.    Heavy 
duty  heaters  may  be  arranged  in  combinations  of  one  or  more  units  in  a 
battery.     A  few  possible  arrangements  are  shown  in  Figs.  6  to  13,  in-, 
elusive. 

Most  manufacturers  of  heavy  duty  furnaces  rate  their  furnaces  in  Btu 
per  hour  and  also  in  the  number  of  square  feet  of  heating  surface.  Con- 
servative practice  indicates  that  at  no  time  in  the  heating-up  period 
should  the  furnace  surface  be  required  to  emit  more  than  an  average  of 
3500  Btu  per  square  foot.  A  higher  rate  of  heat  emission  tends  to  increase 
the  heat  loss  up  the  chimney,  and  raise  fuel  consumption,  to  shorten  the 
life  of  the  furnace,  and  to  overheat  the  air,  The  ratio  of  heating  surface 
to  grate  area  on  furnaces  for  this  type  of  work  should  never  be  less  than 
30  to  1  and  as  indicated  previously  may  run  as  high  as  50  to  1. 

Control  of  temperature  is  secured  through  (1)  controlling  the  quantity 
of  heated  air  entering  the  room,  (2)  using  mixing  dampers,  or  (3)  regu- 
lating the  fuel  supply. 

The  design  of  heavy  duty  fan  furnace  heating  systems  is  in  many 
respects  similar  to  that  of  the  central  fan  heating  systems  described  in 
Chapter  22.  Ducts  are  designed  by  the  method  outlined  in  Chapter  20. 


PROBLEMS  IN  PRACTICE 

1  •  Why  do  furnaces  designed  to  burn  bituminous  coal,  oil,  or  gas  require 
larger  combustion  spaces  than  those  designed  for  anthracite? 

Anthracite  burns  largely  as  fixed  carbon  whereas  gas  and  oil  burn  as  gases,  and  as  much 
as  50  per  cent  of  bituminous  coal  burns  as  a  gas.  Ample  space  must  be  provided  for  the 
intimate  mixture  of  these  gases  with  the  oxygen  of  the  air  to  secure  proper  combustion. 

2  •  A  furnace  has  the  following  dimensions:    Grate  diameter,  24  in.;  casing 
diameter  for  gravity  air  flow,  56  in.;  combustion  chamber  diameter,  30  in. 
What  is  the  unobstructed  area  required  for  passage  of  ah*  across  the  heating 
surface  when  a  motor-driven  hlower,  operating  at  an  outlet  velocity  of  1200  fpm, 
delivers  1600  cfm  into  the  casing  near  its  bottom? 

For  residence  applications  using  small  blowers,  an  air  outlet  velocity  of  about  one  third 
of  the  blower"  outlet  velocity  is  considered  good  practice1. 

387 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

1 200 

Air-pass  velocity  =  — •« —  =  400  fpm. 
o 

Air-pass  area  =  — j—  =  4  sq  ft  =  576  sq  in. 

3  •  In  Question  2  what  would  be  the  gap  between  the  chamber  and  the  baffle 
when  the  chamber  is  centered  in  the  casing? 

Area  of  combustion  chamber  (30-in.  diam)  706.9  sq  in. 

Area  of  air  pass  576.0  sq  in. 

Total  area  1282.9  sq  in. 

The  diameter  of  a  circle  with  an  area  of  1282.9  sq  in.  is  40.4  in.  One  half  of  the  difference 
between  the  diameters  is  the  amount  of  gap. 

Gap  =  — : — H- —  =  5.2  in.  =  approximately  5K  in. 

*  fJ 

4  •  Why  should  secondary  surface  be  designed  for  easy  cleaning? 

If  the  combustion  is  not  perfect,  soot  is  formed  immediately  above  the  fire  and  is  apt 
to  form  a  deposit  on  the  secondary  surface  from  which  it  should  be  removed.  If  the 
secondary  surface  is  so  designed  that  there  are  horizontal  passages,  fine  gray  ash  will 
settle  out  in  these  to  form  an  insulation  between  the  hot  gases  of  combustion  and  the 
metal  of  the  furnace;  consequently,  these  should  be  readily  cleaned.  If  the  passages  are 
vertical  they  are  largely  self-cleaning  of  ash,  but  provision  should  be  made  for  easy  and 
thorough  cleaning  of  the  collection  chamber  below  them. 

5  •  Why  is  baffling  inside  the  casing  necessary  on  fan  systems? 

Because  the  movement  of  air  is  independent  of  its  temperature,  air  must  be  guided  by 
baffles  of  one  form  or  another  to  bring  it  in  contact  with  the  hot  surfaces  so  it  will  not 
pass  through  the  casing  unheated.  On  the  other  hand,  if  the  air  is  held  against  a  hot 
surface  too  long  it  might  become  overheated,  for  the  average  register  temperature  on  a 
fanjsystem  should  not  exceed  120  F. 

6  •  Why  do  buildings  which  are  intermittently   used  require  more  heating 
capacity  than  buildings  constantly  used? 

Between  heating  periods  the  intermittently  used  building  is  allowed  to  cool  down.  All 
of  the  material  in  the  building  loses  heat,  and  before  the  building  can  be  reheated  to  a 
comfortable  temperature  this  material  must  also  be  reheated. 

7  •  What  practical  points  should  be  observed  in  designing  a  fan  system  in  order 
to  eliminate  noise? 

a>.  Use  a  large  fan  so  it  can  be  run  at  slow  speed. 

b.  Set  the  fan  and  motor  on  a  solid  foundation. 

c.  Insulate  the  fan  and  motor  from  the  foundation  with  rubber,  cork,  or  other  springy 
material  according  to  the  principles  given  in  Chapter  18,  provided,  of  course,  that  such 
insulation  is  of  value. 

d.  See  that  the  air  velocity  is  not  too  high  in  the  ducts.    Properly  designed  splitters  in 
the  elbows  will  avoid  high  velocities  at  the  turns  in  cases  where  the  velocity  through  the 
ducts  themselves  is  not  too  high. 

e~  Use  canvas  connections  between  the  ducts  and  any  running  equipment. 
/.   Be  sure  the  ducts  have  a  relatively  smooth  interior  and  are  rigid. 


388 


Chapter  24 

GRAVITY  WARM  AIR  FURNACE 
SYSTEMS 

Procedure  for  Design,  Estimating  Heating  Requirements,  Sizes  of 

Leader  Pipes,  Proportioning  Wall  Stacks,  Register  Sizes,  Recircu- 

lating  Ducts  and  Grilles,  Return  Connection  to  Furnace,  Furnace 

Capacity,    Examples,    Booster    Fans 

WARM  air  heating  systems  of  the  gravity  type  are  described  in  this 
chapter1,  and  those  of  the  mechanical  type  are  described  in  Chapter 
23.  In  the  gravity  type,  the  motive  head  producing  flow  depends  upon 
the  difference  in  weight  between  the  heated  air  leaving  the  top  of  the 
casing  and  the  cooled  air  entering  the  bottom  of  the  casing,  while  in  the 
mechanical  type  a  fan  may  supply  all  or  part  of  the  motive  head.  Booster 
fans  are  often  used  in  conjunction  with  gravity-designed  systems  to 
increase  air  circulation. 

In  general,  a  warm-air  furnace  heating  plant  consists  of  a  fuel-burning 
furnace  or  heater,  enclosed  in  a  casing  of  sheet  metal  or  brick,  which  is 
placed  in  the  basement  of  the  building.  The  heated  air,  taken  from  the 
top  or  sides  near  the  top  of  the  furnace  casing,  is  distributed  to  the 
various  rooms  of  the  building  through  sheet  metal  warm-air  pipes.  The 
warm-air  pipes  in  the  basement  are  known  as  leaders,  and  the  vertical 
warm-air  pipes  which  are  run  in  the  inside  partitions  of  the  building  are 
called  stacks.  The  heated  air  is  finally  discharged  into  the  rooms  through 
registers  which  are  set  in  register  boxes  placed  either  in  the  floor  or  in 
the  side  wall,  usually  at  or  near  the  baseboard. 

The  air  supply  to  the  furnace  may  be  taken  (1)"  entirely  from  inside 
the  building  through  one  or  more  recirculating  ducts,  (2)  entirely  from 
outside  the  building,  in  which  case  no  air  is  recirculated,  or  (3)  through  a 
combination  of  the  inside  and  the  outside  air  supply  systems.  \ 

PROCEDURE  FOR  DESIGN 

The  design  of  a  furnace  heating  system  involves  the  determination 
of  the  following  items: 

1.  Heat  loss  in  Btu  from  each  room  in  the  building. 

2.  Area  and  diameter  in  inches  of  warm-air  pipes  in  basement  (known  as  leaders). 

3.  Area  and  dimensions  in  inches  of  vertical  pipes  (known  as  wall  stacks), 

4.  Free  and  gross  area  and  dimensions  in  inches  of  warm-air  registers, 

5.  Area  and  dimensions  of  recirculating  or  outside  air  ducts,  in  inches. 

6.  Free  and  gross  area  and  dimensions  in  inches  of  recirculating  registers. 


JAn  figures  and  much  of  the  engineering  data  which  follow  are  from  Bulletins  No.  141,  188  and  189, 
Warm  Air  Furnaces  and  Heating  Systems,  Part  II,  by  Professor  A.  C.  Willard,  A.  P.  Kratz,  and  V.  SL 
Day,  Engineering  Experiment  Station,  University  of  Illinois. 

389 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


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CHAPTER  24 — GRAVITY  WARM  AIR  FURNACE  SYSTEMS 

7.  Size  of  furnace  necessary  to  supply  the  warm  air  required  to  overcome  the  heat 
loss  from  the  building.    This  size  should  include  square  inches  of  leader  pipe  area  which 
the  furnace  must  supply.     It  is  also  desirable  to  call  for  a  minimum  bottom  fire-pot 
diameter  in  inches,  which  is  the  nominal  grate  diameter.  • 

8.  Area  and  dimensions  in  inches  of  chimney  and  smoke  pipe.    If  an  unlined  chimney 
is  to  be  used,  that  fact  should  be  made  clear. 

The  heat  loss  calculations  should  be  made  in  accordance  with  the 
procedure  outlined  in  Chapter  7,  taking  into  consideration  the  trans- 
mission losses  as  well  as  the  infiltration  losses/ 

SIZES  OF  LEADER  PIPES 

In  a  gravity  circulating  warm-air  furnace  system  the  size  of  the  leader 
to  a  given  room  depends  upon  the  temperature  of  the  warm  air  entering 
the  room  at  the  register.  A  reasonable  air  temperature  at  the  registers 
must,  therefore,  be  chosen  before  the  system  can  be  designed.  The 
National  Warm  Air  Heating  Association  has  approved  an  air  temperature 
of  175  F  at  the  registers  as  satisfactory  for  design  purposes.  At  this  tem- 
perature, the  heat-carrying  capacity  (heat  available  above  70  F)  per 
square  inch  of  leader  pipe  per  hour  for  first,  second  or  third  floors  is  shown 
by  Fig.  1  at  175  F  to  be  105,  170  and  208  Btu,  respectively.  For  average 
calculations,  the  values  111,  166  and  200  will  simplify  the  work  and  may 
be  satisfactorily  substituted  for  these  heat-carrying  capacities.  If  H 
represents  the  total  heat  to  be  supplied  any  room,  the  resulting  equations 
are: 

Leader  areas  for  first  floor,  square  inches  —  —  =  approximately  0.009IT          (1) 

H 
Leader  areas  for  second  floor,  square  inches  =  r^r  =  approximately  0.006H      (2) 

rr 

Leader  areas  for  third  floor,  square  inches  =  ^rr  »  approximately  0.005J3"        (3) 

In  designing  for  a  lower  warm-air  register  temperature,  say  160  F,  the 
factors  111,  166  and  200  become  80,  140  and  166  (Fig.  1  at  160  F),  and 
the  resulting  equations  are: 

TT 

Leader  areas  for  first  floor,  square  inches  =  -^r-  =  approximately  0.012-ET  (4) 

•rr 

Leader  areas  for  second  floor,  square  inches  =  -rrr  —  approximately  0.0075'     (5) 

Leader  areas  for  third  floor,  square  inches  =   r^  =  approximately  O.OO&H        (6) 

loo 

These  equations  are  applicable  to  straight  leaders  from  6  to  8  ft  in 
length.  Longer  leaders  must  be  very  thoroughly  covered  or  else  the 
vertical  stacks  must  be  increased  in  area  as  discussed  under  wall  stacks. 
If  some  provision  is  not  made  for  these  longer  leaders,  the  air  tempera- 
ture may  be  much  lower  than  anticipated  and  the  room  will  not  be 
properly  heated. 

While  Fig.  1  takes  care  of  the  drop  in  temperature  in  straight  leaders 
up  to  8  ft  in  length  connected  to  stacks  having  about  75  per  cent  of  the 

391 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


area  of  the  leader,  the  designer  must  make  allowances  for  all  other 
conditions.  The  temperature  drop  in  leaders  of  various  lengths  at  three 
different  register  temperatures  is  shown  in  Fig.  2,  and  should  be  used  to 
obtain  new  register  temperatures,  lower  than  175  F,  on  which  to  base 
selections  from  the  curves  of  Fig.  1,  and  thereby  new  constants  for 
Equations  1,  2  and  3. 

Leader  sizes  should  in  general  be  not  less  than  those  obtained  by 
Equations  1  to  3  nor  should  leaders  less  than  8  in.  in  diameter  be  used.  It 
is  not  considered  good  commercial  practice  to  specify  diameters  except 


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FIG.  3.    RELATIVE  HEATING  EFFECT  OF  STACKS  AT  CONSTANT  HEAT 
INPUT  TO  FURNACE 

Note. — Exterior  surface  of  all  ducts  is  bright  tin  except  at  joints 
'      •  where  asbestos  sealing  strips  are  used. 

in  whole  inches.  The  tops  of  all  leaders  should  be  at  the  same  elevation  as 
they  leave  the  furnace  bonnet,  and  from  this  point  there  should  be  a 
uniform  up-grade  of  1  in.  per  foot  of  run  in  all  cases.  Leaders  over  12  ft 
in  length  are  to  be  avoided  or  should  receive  very  special  attention. 

PROPORTIONING  WALL  STACKS 

The  wall  stack  for  an  upper  floor  should  be  made  not  less  than  70 
per  cent  of  the  area  of  the  leader  which  has  been  selected  from  Fig.  1. 
So  long  as  the  leader  is  short  and  straight  as  was  the  case  for  Fig.  1, 

392 


CHAPTER  24 — GRAVITY  WARM  AIR  FURNACE  SYSTEMS 

such  a  practice  is  probably  justified,  since  the  loss  (Fig.  3)  in  capacity 
occasioned  by  the  smaller  stack  is  not  very  serious  for  stacks  having 
areas  in  excess  of  70  per  cent  of  the  leader  area.  For  leaders  over  8  ft 
in  length  or  for  leaders  which  are  not  straight,  the  ratio  of  stack  area  to 
leader  area  should  be  greater  than  70  per  cent  in  order  to  offset  the 
greater  temperature  losses  (Fig.  2)  in  the  longer  leader.  In  gravity 
circulating  systems,  this  stack  to  leader  area  ratio  is  a  very  important 
consideration.  Specific  data  for  a  great  variety  of  cases  are  presented 
in  Figs.  4  and  5  and  the  designer  should  check  the  stack  to  leader  com- 
binations with  the  nearest  comparable  case  as  shown  in  these  figures. 
Any  second-floor  stack  supplying  heat  to  a  room  whose  heat  loss  is  9,000 
Btu  or  more  (see  Figs.  4  and  5  which  show  that  high  temperatures 
are  necessary  if  rooms  of  more  than  9,000  Btu  requirement  are  heated  by 
one  stack  each  in  4-in.  studding)  should  be  run  within  6-in.  studded  walls 
or  should  have  multiple  stacks.  Stack  sections,  wherever  possible,  should 
be  changed  from  the  thin  rectangular  to  the  more  nearly  square  shape. 

REGISTER  SIZES 

The  registers  used  for  discharging  warm  air  into  the  rooms  should  have 
free  or  net  area  not  less  than  the  area  of  the  leader  in  the  same  run  of 
piping.  The  free  area  should  be  at  least  70  per  cent  of  the  gross  area 
of  the  register.  No  upper-floor  register  should  be  wider  horizontally 
than  the  wall  stack,  and  it  should  be  placed  either  in  the  baseboard  or 
side  wall,  if  this  can  be  done  without  the  use  of  offsets.  First-floor  registers 
may  be  of  the  baseboard  or  floor  type,  with  the  former  location  preferred. 

RECIRCULATING  DUCTS  AND  GRILLES 

The  ducts  through  which  air  is  returned  to  the  furnace  should  be 
designed  to  minimize  friction  and  turbulence.  They  should  be  of  ample 
area,  in  excess  of  the  total  area  of  warm-air  pipes,  and  at  all  points  where 
the  air  stream  must  change  direction  or  shape,  streamline  fittings  should 
be  employed.  Horizontal  ducts  should  pitch  at  least  J^  in.  per  foot 
upward  from  the  furnace. 

The  recirculating  grilles  (or  registers)  should  have  a  free  area  at  least 
equal  to  the  ducts  to  which  they  connect,  and  their  free  area  should 
never  be  less  than  50  per  cent  of  their  gross  area. 

The  location  and  number  of  return  grilles  will  depend  on  the  size,  details 
and  exposure  of  the  house.  Small  compactly  built  houses  may  frequently 
be  adequately  served  by  a  single  return  effectively  placed  in  a  central  hall. 
More  often  it  is  desirable  to  have  two  or  more  returns,  provided,  however, 
that  in  two-story  residences  one  return  must  be  placed  to  effectively 
receive  the  cold  air  returning  by  way  of  the  stairs. 

Where  a  divided  system  of  two  or  more  returns  is  used,  the  grilles 
must  be  placed  to  serve  the  maximum  area  of  cold  wall  or  windows. 
Thus  in  rooms  having  only  small  windows  the  grille  should  be  brought 
as  close  to  the  furnace  as  possible,  but  if  the  room  has  a  bay  window, 
French  doors,  or  other  large  sources  of  cooling  or  leakage  of  cold  air,  the 
grille  should  be  placed  close  by,  so  as  to  collect  the  cool  air  and  prevent 
drafts.  When  long  ducts  of  this  type  are  employed  they  must  be  made 

393 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


oversize  and  favored  in  every  way.     This  precaution  is  particularly 
important  when  long  ducts  and  short  ducts  are  used  in  the  same  system. 


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parallel  with  short  ducts. 

Return  ducts  from  upstairs  rooms  may  be  necessary  in  apartments 
or  other  spaces  closed  off  or  badly  exposed.  Metal  linings  are  advisable 
in  such  ducts.  It  is  important  that  these  ducts  be  free  from  unnecessary 

394 


CHAPTER  24 — GRAVITY  WARM  AIR  FURNACE  SYSTEMS 


friction  and  turbulence,  and  that  they  be  located  to  prevent  preheating 
of  the  air  before  it  reaches  the  furnace. 


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Return  Connection  to  Furnace 

'  Circulation  is  accelerated  if  the  drop  to  the  furnace  is  through  a  round 
inclined  pipe  with,  say,  two  45-deg  elbows  rather  than  through  a  vertical 
drop  and  two  90-deg  elbows.  The  top  of  the  shoe  should  never  enter 

395 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


the  casing  above  the  level  of  the  grate  in  the  furnace.    To  accomplish 
this  the  shoe  must  be  wide. 

Tests  of  six  different  systems  of  cold  air  returns,  Fig.  6,  made  at  the 
University  of  Illinois2,  resulted  in  the  following  conclusions: 

1.  In  general,  somewhat  better  room  temperature  conditions  may  be  obtained  by 
returning  the  air  from  positions  near  the  cold  walls. 

2.  Friction  and  turbulence  in  elaborate  return  duct  systems  retard  the  flow  of  air, 
and  may  seriously  reduce  furnace  efficiency,  and  lessen  the  advantages  of  such  a  design. 

3.  The  cross-sectional  duct  area  is  not  the  only  measure  of  effectiveness.    Friction 
and  turbulence  may  operate  to  make  the  air  flow  out  of  all  proportion  to  the  various 
duct  areas. 


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FIG.  6.  ARRANGEMENT  OF  COLD  AIR  RETURNS  FOR  Six  INSTALLATIONS 

FURNACE  CAPACITY 

The  size  of  furnace  should,  of  course,  be  such  as  will  provide  the 
necessary  air  heating  capacity,  usually  expressed  in  square  inches  of 
leader  pipe  area,  and  at  the  same  time  provide  a  grate  of  the  proper 
area  to  burn  the  necessary  fuel  at  a  reasonable  chimney  draft.  The  total 
leader  pipe  area  required  is  easily  obtained  by  finding  the  sum  of  the 
leader  pipe  areas  as  already  designated. 

The  grate  area  will  depend  on  several  factors  of  which  four  are  very 
important.  First  of  all,  the  air  temperature  at  the  register  for  which 
the  plant  has  been  designed  must  be  determined.  Usually,  this  tempera- 
ture is  taken  as  175  F.  Second  in  importance  is  the  combustion  rate, 
which  must  always  correspond  with  the  register  air  temperature,  as  is  shown 
by  reference  to  a  set  of  typical  furnace  performance  curves  (Fig.  7)  for  a 
cast-iron  circular  radiator  furnace  with  a  23-in.  diameter  grate  and  50-in. 
diameter  casing.  The  conditions  shown  on  these  curves  which  seem  to 


a  Investigation  of  Warm-Air  Furnaces  and  Heating  Systems,  Part  N,  by  A.  C.  Willard,  A.  P.  Kratz  and 
V.  S.  Day  (University  of  Illinois  Engineering  Experiment  Station  Bulletin  No.  189). 


CHAPTER  24 — GRAVITY  WARM  AIR  FURNACE  SYSTEMS 

approximate  nearest  to  the  175  F  register  warm-air  temperature  are: 
combustion  rate,  7  Ib;  warm-air  register  temperature,  173  F;  efficiency  of 
the  furnace,  58.5  per  cent.  The  third  factor  is  efficiency,  which,  in  turn,  is 
a  function  of  the  combustion  rate  varying  with  it  as  shown  by  the  effi- 
ciency curve  of  Fig.  7.  The  fourth  factor  is  the  heat  value  per  pound  of 
fuel  burned,  which  was  12,790  Btu.  This  is  not  shown  on  the  curves  since 
it  was  constant  for  all  combustion  rates. 


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2  . 

Comtiusf/0/?  fate  f/7  /h  per  s$  ft  of  Grate  per  tir '  ^ 

FIG.  7.     TYPICAL  PERFORMANCE  CURVES  FOR  A  WARM  AIR  FURNACE  AND  INSTALLATION 
IN  A  THREE  STORY  TEN  LEADER  PLANT,  OPERATING  ON  RECIRCULATED  AIR 

From  the  relation  existing  among  these  factors  it  is  found  (Fig.  7)  that 
the  capacity  of  the  furnace  under  test  is  147,750  Btu  per  hour  for  the  total 
grate,  which  gives  the  capacity  at  the  furnace  bonnet  per  square  foot  of 
grate  as  51,200  Btu  and  per  square  inch  of  grate  as  356  Btu  per  hour. 

Suppose  it  is  desired  to  select  a  furnace  to  deliver  air  to  the  rooms  at  a 
register  temperature  approximating  160  F  rather  than  175  F.  Referring 
to  the  curves,  the  relation  is:  combustion  rate,  5.5  Ib;  register  warm-air 
temperature,  160  F ;  and  efficiency  of  the  furnace,  62  per  cent.  Under  this 
condition  the  capacity  of  the  furnace  at  the  furnace  bonnet  per  square 
foot  of  grate  area  is  43,200  Btu  per  hour,  and  per  square  inch  of  grate  it  is 
300  Btu  per  hour.  From  these  performance  values^  the  grate  area  for  any 

397 


AMERICAN  SOCIETY  of  HEATING  dnJ  VENTILATING  ENGINEERS  GUIDE,  1935 


FIG.  8.   BASEMENT  PLAN,  RESEARCH  RESIDENCE 


FIG,  9.  FIRST-FLOOR  PLAN,  RESEARCH  RESIDENCE 
398 


CHAPTER  24 — GRAVITY  WARM  AIR  FURNACE  SYSTEMS 

plant  requirement  (allowing  20  per  cent  heat  loss  between  furnace  and 
registers)  will  be: 

1  2  H 
Grate  area  (175  F  register  temperature),  square  inches  =  •• '          =  0.0034H"       (7) 

oOD 

Grate  area    (160  F),  square  inches    -   ^^  =0.00405"  (8) 

oUU 

Here  H  =  Btu  heat  loss  from  the  entire  house  per  hour  =  summation 
of  all  room  losses  H i  +  H%  +  etc.  +  the  Btu  necessary  to  heat  the  fresh 
air,  if  any,  at  intake.  This  fresh  air  loss  in  Btu  per  hour  will  be  approxi- 
mately 1.27  times  the  cubic  feet  of  air  admitted  through  the  intake  per 
hour  on  a  zero  day.  For  systems  which  recirculate  all  the  air  this  value 
will  be  zero.  For  systems  which  have  a  fresh  air  intake,  controlled  by 
damper,  this  value  might  well  be  approximated,  since  this  loss  will 
probably  be  reduced  to  a  minimum  on  a  zero  day.  Assume  for  such  cases 
that  the  building  loss  is  increased  by  25  per  cent,  and  that  there  is  the 
usual  20  per  cent  loss  between  furnace  and  registers. 

It  is  not  always  possible  to  obtain  performance  curves,  and  the  fol- 
lowing method  is  suggested  as  being  a  close  check.  An  addition  of  2  per 
cent  of  the  furnace  capacity  is  proposed  for  each  unit  that  the  heating 
surface  to  grate  area  ratio  of  the  furnace  exceeds  20.  This  addition  is 
based  on  tests  made  at  the  University  of  Illinois,  of  four  types  of  furnaces 
having  various  ratios  of  heating  surface  to  grate  area. 

Let  E  «  efficiency  of  the  furnace. 

/  =  fuel  value  of  the  coal,  Btu  per  pound. 

p  =  pounds  of  coal  burned  per  square  foot  of  grate  per  hour. 

R  =  ratio  of  heating  surface  to  grate  area. 

H  ~  total  heat  requirements  of  the  house. 

1  2  X  111  H 
Grate  area,  square  inches  =  £ .    f  1  +  Q  02  (.R  ~  20)  ]  f°r  al*  insi(*e  ain          ^ 

For  coal  having  a  heat  value  of  12,000  Btu,  and  a  furnace  having  60  per 
cent  efficiency,  with  6  Ib  of  coal  burned  per  square  foot  of  grate  per  hour, 
and  20  sq  ft  of  heating  surface  for  1  sq  ft  of  grate,  this  becomes : 

1  2  X  144  H 
Grate  area,  square  inches  *  Q  60 '  X  12  000  X  6  f°f  a11  insi<le  air' 

and  for  another  furnace  having  24  sq  ft  of  heating  surface  for  1  sq  ft 
of  grate  the  expression  is 

.    .  1.2  X  144  H ,in 

Grate  area,  square  inches  =  0.60  X  12,000  X  6  [1+ 0.02  (24  ~- 20)]  (11) 

The  air  temperatures  at  the  registers  corresponding  to  the  conditions 
of  Equation  11  would  be  approximately  165  F,  and  for  175  F  and  12,000 
Btu  the  combustion  rate  would  be  about  7.5  Ib  with  an  efficiency  of 
57  per  cent,  using  the  curves  of  Fig.  7  as  a  guide. 

TYPICAL  DESIGN 

The  application  of  the  preceding  data  to  an  actual  example  may  be  of 
assistance  to  the  designer.  Figs.  8,  9,  10  and  11  represent  the  plans  of 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


FIG.  10.   SECOND-FLOOR  PLAN,  RESEARCH  RESIDENCE 


FIG.  11.   THIRD-FLOOR  PLAN,  RESEARCH  RESIDENCE 


'400 


CHAPTER  24 — GRAVITY  WARM  AIR  FURNACE  SYSTEMS 

the  Warm  Air  Research  Residence  of  the  National  Warm  Air  Heating 
Association  erected  at  the  University  of  Illinois3. 

Leaders,  Stacks  and  Registers.    (Direct  Method) 

Living  Room,  1st  floor: 

17,250  -r-  111  =  155  sq  in.  leader  area.   See  summary,  Table  1;  also  example  under 
Standard  Code4,  Art.  3,  Basis  of  Working  Rules  for  Pipes. 

Leader  diameter  =  14  in. 

Register  size  =  155  sq  in.  net  area.    Gross  area  =  net  area  -f-  0.7  =  14  in.  X  16  in. 

Owner's  Room,  2nd  floor: 

15,030  -r-  167  =  90  sq  in.  leader  area.     See  Summary  Table;  also   example  under 
Standard  Code4,  Art.  3,  Basis  of  Working  Rules  for  Pipes. 

Leader  diameter  =  11.4,  say  12  in. 

Stack  area  =  0.7  X  90  =  63  sq  in.  =  say  5  in.  X  12  in. 

Register  area        =  90  sq  in.  net  area.     Gross  area  =  net  area  -f-  0.7  =  12  X  12 
or  12  in.  X  14  in. 

In  like  manner  the  leaders,  stacks  and  registers  are  calculated  for  each 
room  in  the  house. 

Leaders,  Stacks  and  Registers.     (Code  4  Method.    See  Art.  3,  Sec.  1,  2,  3.) 

Living  Room  (Glass  -  90,  Net  wall  =  405,  Cubic  contents  =  2405) 

T      -  /  90     .    405    ,     2405  \  n       1KK       . 

Leader  =    (^_  +  —  +  w  J  9  =  155  sq  m. 

Register,  same  as  Direct  Method. 

Owner's  Room  (Glass  =  68,  Net  wall  =  394,  Cubic  contents  =  2275) 

T      .  /  68    ,    394    ,     2275 

Leader  -    ^_+_  + 

*  Stack  and  Register,  same  as  Direct  Method. 

Assuming  all  air  recirculated,   the  minimum  furnace  for  the  plant 
will  be : 

Grate  area  =  0.0034  X  132,370  =  450  sq  in.  «  24  in.  diameter  at  175  F 
register  temperature.     (Equation  7) 

Grate  area  «  0.0040  X  132,370  «  530  sq  in.  =  26  in.  diameter  at  160  F 
register  temperature.     (Equation  8) 

If  provision  should  be  made  for  certain  outside  air  circulation,  then 
increase  the  building  heat  loss  by,  say  25  per  cent  and  obtain  by  Equation 
7  a  27-in.  grate  and  by  Equations  8  and  10  a  29-in.  grate. 

Experiments  at  the  University  of  Illinois5  have  shown  that  the  capacity 
of  a  furnace  may  be  increased  nearly  three  times  by  an  adequate  fan, 


3Plans  used  with  permission.    Bathroom  on  third  floor  not  heated. 

^Standard  Code  Regulating  the  Installation  of  Gravity  Warm  Air  Heating  Systems  in  Residences. 
This  code  has  been  sponsored  by  the  National  Warm  Air  Heating  Association,  the  National  Association  of 
Sheet  Metal  Contractors,  and  the  AMERICAN  SOCIETY  OF  HEATING  AND  VENTILATING  ENGINEERS.  It  is 
recommended  that  the  installation  of  all  gravity  warm  air  heating  systems  in  residences  be  governed  by 
the  provisions  of  this  code,  the  eighth  edition  of  which  may  be  obtained  from  the  National  Warm  Air  Heating 
Association,  3440  A.I.U.  Building,  Columbus,  Ohio. 

*See  University  of  Illinois  Eng.  Exp.  Sla.  Bulletin  No.  120,  p.  129. 

401 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


with  a  constant  register  or  delivery  temperature  maintained,  provided 
that  the  rate  of  fuel  consumption  can  be  increased  to  provide  the  necessary 
heat.  In  other  words,  the  capacity  of  a  forced  circulation  systern  is  limited 
by  the  ability  of  the  chimney  to  produce  a  sufficient  draft. 

TABLE  1.    SUMMARY  OF  DATA  APPLIED  TO  WARM  AIR  RESEARCH  RESIDENCE 


Rooms 

From 
Chapter  7 
Estimating 
Heat  Losses 
Btu 
Heat  Losses 
H 

Leader 
Area 
SQ  In. 

Stack  Area 
Sq   In. 
0.7  X  LA 

Leader 
Diameter 

Inches 

Stack 
Size 

Net 

Register 
Size 
Gross 

First  Floor 
Livinsr  . 

17250 

«  0.0091? 
155 

14 

14  X  16 

AwA  V  lllg  

Dining—.  

Kitchen  
Sun 

6810 
2300 
9210 
25710 

61 
21 
83 
230 

— 

9 
8 
11  or  12 
Two  12 



8X12 
8  X  10 
12  X14 
Two  12  X  14 

Hall  and  stair 
Second  Floor 
Owner's  
S.  W.  Bed...... 
Bath  
N.  Bed  

12570 

15030 
9800 
2450 
14800 

113 
»  0.006H 
90 
59 
15 
89 

63 
41 
10 
62 

12 

11  or  12 
9 
8 
11  or  12 

5X  12 
3J^X  12 
3X  10 
5  X  12 

12  X  14 

12  X  14 
8X12 
8  X  10 
12  X  14 

Third  Floor 
E.  Bed  
W.  Bed  

8220 
8220 

-  0.005H 
41 
41 

29 
29 

8 
8 

3X  10 
3  X  10 

8X10 
8  X  10 

BOOSTER  FANS 

Booster  fans  often  may  be  arranged  to  operate  when  gas  or  oil  burners 
are  running  and  to  stop  automatically  when  the  burners  shut  down.  The 
booster  equipment  is  most  effective  in  increasing  output  at  low  operating 
temperatures.  According  to  tests,  efficiencies  may  be  advanced  from  60 
per  cent  for  gravity  to  70  per  cent  with  boosters  at  low  operating  tem- 
peratures, but  at  high  operating  temperatures  gravity  and  booster 
efficiencies  are  almost  identical6. 


•See  University  of  Illinois  Bng.  Exp.  Sta.  Bulletin  No.  141,  p.  79. 


PROBLEMS  IN  PRACTICE 

1  •  A  facet  story  dining  room  has  a  calculated  heat  loss  of  12,000  Btu  per  hour. 

a.  What  size  leader  pipe  should  he  used  for  175  F  register  air  temperature? 

b.  "What  size  register? 

12  000 
a.  Leader  area  =*  ~~TTi —  ~  108.1  sq  in.    Use  leader  with  diameter  of  12  in. 


1).  Register  gross  area 


108 

•JT-=  —  154  sq  in.  Use  12  in.  by  14  in.  register. 

402 


CHAPTER  24  —  GRAVITY  WARM  AIR  FURNACE  SYSTEMS 

2  •  A  third-story  bedroom  has  a  calculated  heat  loss  of  12,000  Btu  per  hour. 

a.  What  size  leader  pipe  should  be  used  for  a  175  F  register  air  temperature? 

b.  What  size  stack? 

c.  What  size  register? 

a.  Leader  area  =       '         =  60  sq  in.    Use  leader  with  diameter  of  9  in. 

•4UU 

b.  Stack  area  =  0.7  X  60  =  42  sq  in.    Use  stack  %y%  in.  by  12  in. 

60 

c.  Register  gross  area  —  —  =  85.7  sq  in.     Use  register  8  in.  by  12  in. 

3  •  The  calculated  heat  loss  of  a  house  is  130,000  Btu  per  hour.    Find  the  grate 
area  required  for  the  furnace  under  the  following  conditions  : 

Heating  value  of  coal  =  12,500  Btu  per  Ib. 

Furnace  efficiency  =  55  per  cent. 

Combustion  rate  —  7.5  Ib  per  sq  ft  per  hour. 

Ratio  of  heating  surface  to  grate  area  of  furnace  =  20  to  1. 

Register  temperature  —  175  F. 

Loss  between  furnace  and  registers  =  20  per  cent. 
-    ^  1.2  X  144  X  130,000       onn  _       . 

Gratearea  =  0.60  X  12,500  XTS  ~  3"'5  Sq  m' 
Grate  diameter  =  22.6  in. 
Use  grate  with  diameter  of  23  in. 

4  •  If  in  Question  3  the  conditions  were  the  same  except  that  the  ratio  of 
heating  surface  to  grate  area  of  furnace  was  24  to  1,  what  size  grate  would  be 
required  for  the  furnace? 

_.  1.2  X  144  X  130,000  ..  1  399.5       Q7n       . 

Gratearea  »  12>500  x  7>5  X    x  +  0.02  (24  -  20)    =  T08    =  370  sq  m. 


Grate  diameter  =  21.7  sq  in. 
Select  grate  with  diameter  of  22  in. 

5  •  Name  the  items  involved  in  the  design  of  a  furnace  heating  system. 

a.  Heat  loss  from  each  room,  Btu. 

b.  Area  and  dimensions  of  warm-air  pipes  in  basement,  inches. 

c.  Area  and  dimensions  of  vertical  pipes,  inches. 

d.  Free  and  gross  area  and  dimensions  of  warm-air  registers,  inches. 

e.  Area  and  dimensions  of  recirculating  or  outside  air  ducts,  inches. 

/.    Free  and  gross  area  and  dimensions  of  recirculating  registers,  inches. 

g.  Size  of  furnace  necessary  to  supply  the  warm  air  to  overcome  the  heat  loss. 

h.  Area  and  dimension  of  chimney  and  smoke  pipe,  inches. 

6  •  Discuss  the  design  features  of  recirculating  ducts. 

a.  Their  area  should  be  equal  to  or  greater  than  that  of  the  supply  ducts. 

b.  They  should  be  streamlined,  and  have  a  minimum  number  of  turns. 

c.  All  runs  should  be  as  short  as  possible. 

d.  Account  should  be  taken  of  all  cold  walls  and  window  areas  in  determining  sizes  and 
positions  of  return  air  inlets. 

e.  The  return  line  should  be  pitched  downward  toward  the  furnace.     It  should  be 
designed  to  minimize  friction. 

/.   The  top  of  the  shoe  or  boot  should  never  be  above  the  grate  level. 

403 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

7  •  Discuss  the  use  of  a  booster  fan.  What  effect  has  a  booster  fan  at  low 
operating  temperatures?  At  high  ones? 

A  booster  fan  is  useful  in  accelerating  the  air  flow  past  the  surface  of  a  low  temperature 
furnace,  where  only  a  small  weight  differential  in  the  air  is  created,  and  in  unbalancing 
a  gravity  system  so  flow  is  established.  The  first  use  involves  the  entire  plant,  and 
increases  efficiency  about  10  per  cent  with  low  temperature  operation;  the  second 
involves  only  the  leaders  in  which  air  flow  is  accelerated.  At  high  operating ^ tempera- 
tures the  difference  in  weight  between  warm  outgoing  air  and  cool  incoming  air  is  great 
enough  to  make  a  booster  unnecessary  with  ordinary  gravity  systems. 


404 


Chapter  25 

BOILERS 

Cast-Iron  Boilers,  Steel  Boilers,  Special  Heating  Boilers,  Gas- 
Fired  Boilers,  Hot  Water  Supply  Boilers,  Furnace  Design,  Heating 
Surface,  Testing  and  Rating  Codes,  Output,  Efficiency,  Selection 
of  Boilers,  Connections  and  Fittings,  Erection,  Operation  and 
Maintenance,  Boiler  Insulation 

STEAM  and  hot  water  boilers  for  low  pressure  heating  work  are  built  in 
a  wide  variety  of  types,  many  of  which  are  illustrated  in  the  Catalog 
Data  Section,  and  are  classified  as  (1)  cast-iron  sectional,  (2)  steel  fire 
tube,  (3)  steel  water  tube,  and  (4)  special. 

CAST-IRON  BOILERS 

Cast-iron  boilers  may  be  of  round  pattern  with  circular  grate  and  hori- 
zontal pancake  sections  joined  by  push  nipples  and  tie  rods,  or  of  rec- 
tangular pattern  with  vertical  sections.  The  latter  type  may  be  either  of 
outside  header  construction  where  each  section  is  independent  of  the  other 
and  the  water  and  steam  connections  are  made  externally  through  these 
headers,  or  assembled  with  push  nipples  and  tie  rods,  in  which  case  the 
water  and  steam  connections  are  internal. 

Cast-iron  boilers  usually  are  shipped  knocked  down  to  facilitate  hand- 
ling at  the  place  of  installation  where  assembly  is  made.  One  of  the  chief 
advantages  of  cast-iron  boilers  is  that  the  separate  sections  can  be  taken 
into  or  out  of  basements  and  other  places  more  or  less  inaccessible  after 
the  building  is  constructed.  This  feature  is  of  importance  in  making 
repairs  to  or  replacing  a  damaged  or  worn  out  boiler  and  should  be  given 
consideration  in  the  original  selection.  Sufficient  space  should  be  pro- 
vided in  the  boiler  room  for  assembling  the  boiler  and  for  disassembling  it 
conveniently  if  repairs  are  needed.  With  the  outside  header  type  of  boiler 
a  damaged  section  in  the  middle  of  the  boiler  can  be  removed  without 
disturbing  the  other  sections  and  sufficient  side  clearance  should  be 
provided  for  this  contingency. 

Capacities  of  cast-iron  boilers  range  from  that  required  for  small 
residences  up  to  about  18,000  sq  ft  of  steam  radiation.  For  larger  loads, 
cast-iron  boilers  must  be  installed  in  multiple,  or  a  steel  boiler  must  be 
used.  In  most  cases  cast-iron  boilers  are  limited  to  working  pressures  of 
15  Ib  for  steam  and  30  Ib  for  water.  Special  types  are  built  for  hot  water 
supply  which  will  withstand  higher  local  water  pressures. 

STEEL  BOILERS 

Two  general  classifications  may  be  applied  to  steel  boilers:  first,  with 
regard  to  the  relative  position  of  water  and  hot  gases,  distinguished  as  fire 

405 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


tube  or  water  tube;  second,  with  regard  to  arrangement  of  furnace  and 
flues,  as  (1)  horizontal  return  tubular  (HRT)  boilers,  (2)  portable  (self- 
contained)  firebox  boilers  with  either  water  or  fire  tubes,  and  (3)  water 
tube  boilers  of  the  power  type. 

Fire  tube  boilers  are  constructed  so  that  the  water  available  to  produce 
steam  is  contained  in  comparatively  large  bodies  distributed  outside  of  the 
boiler  tubes,  the  hot  gases  passing  within  the  tubes.  In  water  tube  boilers, 
the  water  is  circulated  within  the  boiler  tubes,  heat  being  applied  ex- 
ternally to  them. 

The  HRT  boiler  is  the  oldest  type  and  consists  of  a  horizontal  cylin- 
drical shell  with  fire  tubes,  enclosed  in  brickwork  to  form  the  furnace  and 

TABLE  1.    PRACTICAL  COMBUSTION  RATES  FOR  SMALL  COAL-FIRED  HEATING  BOILERS 
OPERATING  ON  NATURAL  DRAFT  OF  FROM  J4  IN.  TO  y%  IN.  WATER* 


KIND  OP  COAL 

SQ  FT  GRATE 

LB  OP  COAL  PER  SQ  FT 
GRATE  PER  HOUR 

No.  1  Buckwheat  Anthracite 

Up  to  4 
5  to  9 
10  to  14 
15  to  19 
20  to  25 

3 

3H 

4 

4H 
5 

Anthracite  Pea 

Up  to  9 
10  to  19 
20  to  25 

5 
5H 
6 

Anthracite  Nut  and  Larger 

Up  to  4 
5  to  9 
10  to  14 
15  to  19 
20  to  25 

8 
9 
10 
11 
13 

Bituminous 

Up  to  4 
5  to  14 
15  and  above 

9.5 
12 
15.5 

aSteel  boilers  usually  have  higher  combustion  rates  for  grate  areas  exceeding  15  sq  ft  than  those  indicated 
in  this  table. 

combustion  chamber.  All  heating  surfaces  and  the  interior  of  the  boiler 
are  accessible  for  both  cleaning  and  inspection.  Horizontal  return  tubular 
boilers,  especially  the  larger  sizes,  should  be  suspended  from  structural 
columns  and  beams  independent  of  the  brick  setting.  Small  HRT  boilers 
sometimes  are  supported  by  brackets  resting  on  the  brick  setting. 

Portable  firebox  boilers  are  the  more  generally  used  type  of  steel  heating 
boilers,  their  outstanding  characteristic  being  the  water- jacketed  firebo?c 
which  eliminates  virtually  all  brickwork.  They  are  shipped  in  one  piece 
from  the  factory  and  come  to  the  job  ready  for  immediate  hook-up  to 
piping.  They  may  be  of  welded  or  riveted  construction  and  have  either 
water  or  fire  tubes.  Manufacturers'  catalogs  usually  list  heating  surface 
as  well  as  grate  area.  The  elimination  of  brickwork  also  makes  this  type 
the  most  compact  of  steel  boilers  as  well  as  the  lowest  in  first  cost. 

Water  tube  boilers.  For  large  heating  loads  water  tube  boilers  are  quite 
frequently  used.  They  usually  require  more  head  room  than  other  types 
of  boilers  but  require  considerably  less  floor  space  and  make  possible  a 

406 


CHAPTER  25 — BOILERS 


much  higher  rate  of  evaporation  per  square  foot  of  heating  surface,  with 
proper  setting,  baffling  and  draft.  Water  tube  boilers  used  for  heating 
purposes  are  brick  set,  supported  on  structural  steel  columns  and  have  the 
brick  setting  encased  in  an  insulated  steel  housing  to  prevent  air  infiltra- 
tion and  to  minimize  heat  losses.  For  large  heating  loads  at  a  high  rate  of 
evaporation,  such  boilers  should  be  operated  at  pressures  above  15  Ib  per 
square  inch  with  a  pressure-reducing  valve  on  the  connection  to  the 
heating  main. 

SPECIAL  HEATING  BOILERS 

A  special  type  of  boiler,  known  as  the  magazine  feed  boiler,  has  been 
developed  for  the  burning  of  small  sizes  of  anthracite.  These  are  built  of 
both  cast-iron  and  steel,  and  have  a  large  fuel  carrying  capacity  which 
results  in  longer  firing  periods  than  would  be  the  case  with  the  standard 
types  using  buckwheat  sizes  of  coal.  Special  attention  must  be  given  to 
insure  adequate  draft  and  proper  chimney  sizes  and  connections. 

Oil-burner  boiler  units,  in  which  a  special  boiler  has  been  designed  with 
a  furnace  shaped  to  suit  the  particular  burner  used,  have  been  developed 
by  a  number  of  manufacturers.  These  usually  are  compact  units  with  the 
burner  and  all  controls  enclosed  within  an  insulated  steel  jacket.  Ample 
furnace  volume  is  provided  for  efficient  combustion,  and  the  heating 
surfaces  are  proportioned  for  effective  heat  transfer.  Consequently, 
higher  efficiencies  are  obtainable  than  with  the  ordinary  coal  fired  boiler 
converted  to  oil  firing. 

GAS-FIRED  BOILERS 

Gas  boilers  have  assumed  a  well-defined  individuality.  The  usual  boiler 
is  sectional  in  construction  with  a  number  of  independent  burners  placed 
beneath  the  sections.  In  most  boilers  each  section  has  its  own  burner.  In 
all  cases  the  sections  are  placed  quite  closely  together,  much  closer  than 
would  be  possible  when  burning  a  soot-forming  fuel.  The  effort  of  the 
designer  is  always  to  break  the  hot  gas  up  into  thin  streams,  so  that  all 
particles  of  the  heat-carrying  gases  can  come  as  close  as  possible  to  the 
heat-absorbing  surfaces.  Because  there  is  no  fuel  bed  resistance  and  because 
the  gas  company  supplies  the  motive  power  to  draw  in  the  air  necessary 
for  combustion  (in  the  form  of  the  initial  gas  pressure),  draft  losses  through 
gas  boilers  are  low. 

HOT  WATER  SUPPLY  BOILERS 

Boilers  for  hot  water  supply  are  classified  as  direct,  if  the  water  heated 
passes  through  the  boiler,  and  as  indirect,  if  the  water  heated  does  not 
come  in  contact  with  the  water  or  steam  in  the  boiler. 

Direct  heaters  are  built  to  operate  at  the  pressures  found  in  city  supply 
mains  and  are  tested  at  pressures  from  200  to  300  Ib  per  square  inch. 
The  life  of  direct  heaters  depends  almost  entirely  on  the  scale-making 
properties  of  the  water  supplied.  If  water  temperatures  are  maintained 
below  140  F  the  life  of  the  heater  will  be  much  longer  than  if  higher 
temperatures  are  used,  owing  to  decreased  scale  formation  and  minimized 
corrosion  below  140  F.  Direct  water  heaters  in  some  cases  are  designed 
to  burn  refuse  and  garbage. 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

Indirect  heaters  generally  consist  of  steam  boilers  in  connection  with 
heat  exchangers  of  the  coil  or  tube  types  which  transmit  the  heat  from  the 
steam  to  the  water.  This  type  of  installation  has  the  following  advantages : 

1.  The  boiler  operates  at  low  pressure. 

2.  The  boiler  is  protected  from  scale  and  corrosion, 

3.  The  scale  is  formed  in  the  heat  exchanger  in  which  the  parts  to  which  the  scale 
is  attached  can  be  cleaned  or  replaced.    The  accumulation  of  scale  does  not  affect 
efficiency  although  it  will  affect  the  capacity  of  the  heat  exchanger. 

4.  Discoloration  of  water  may  be  prevented  if  the  water  supply  comes  in  contact 
with  only  non-ferrous  metal. 

Where  a  steam  heating  system  is  installed,  the  domestic  hot  water 
usually  is  obtained  from  an  indirect  heater  placed  below  the  water  line  of 
the  boiler. 

FURNACE  DESIGN 

Good  efficiency  and  proper  boiler  performance  are  dependent  on  cor- 
rect furnace  design  embodying  sufficient  volume  for  burning  the  par- 
ticular fuel  at  hand,  which  requires  thorough  mixing  of  air  and  gases  at 
a  high  temperature  with  a  velocity  low  enough  to  permit  complete  com- 
bustion of  all  the  volatiles.  On  account  of  the  small  amount  of  volatiles 
contained  in  coke,  anthracite,  and  semi-bituminous  coal,  these  fuels  can 
be  burned  efficiently  with  less  furnace  volume  than  is  required  for  bi- 
tuminous coal,  the  combustion  space  being  proportioned  according  to  the 
amount  of  volatiles  present. 

Combustion  should  take  place  before  the  gases  are  cooled  by  the  boiler 
heating  surface,  and  the  volume  of  the  furnace  must  be  sufficient  for  this 
purpose.  The  furnace  temperature  must  be  maintained  sufficiently  high 
to  produce  complete  combustion,  thus  resulting  in  a  higher  CO 2  content 
and  the  absence  of  CO.  Hydrocarbon  gases  ignite  at  temperatures 
varying  from  1000  to  1500  F. 

The  question  of  furnace  proportions,  particularly  in  regard  to  mechani- 
cal stoker  installations,  has  been  given  some  consideration  by  various 
manufacturers'  associations.  Arbitrary  values  have  been  recommended 
for  minimum  dimensions.  A  customary  rule-of-thumb  method  of  figuring 
furnace  volumes  is  to  allow  1  cu  ft  of  space  for  a  maximum  heat  release 
of  50,000  Btu  per  hour.  This  value  is  equivalent  to  allowing  approxi- 
mately 1  cu  ft  for  each  developed  horsepower,  and  it  is  approved  by 
most  smoke  prevention  organizations. 

The  setting  height  will  vary  with  the  type  of  stoker.  In  an  overfeed 
stoker,  for  instance,  all  the  volatiles  must  be  burned  in  the  combustion 
chamber  and,  therefore,  a  greater  distance  should  be  allowed  than  for  an 
underfeed  stoker  where  a  considerable  portion  of  the  gas  is  burned  while 
passing  through  the  incandescent  fuel  bed.  The  design  of  the  boiler  also 
may  affect  the  setting  height,  since  in  certain  types  the  gas  enters  the 
tubes  immediately  after  leaving  the  combustion  chamber,  while  in  others 
it  passes  over  a  bridge  wall  and  toward  the  rear,  thus  giving  a  better 
opportunity  for  combustion  by  obtaining  a  longer  travel  before  entering 
the  tubes. 

To  secure  suitable  furnace  volume,  especially  for  mechanical  stokers  or 
oil  burners,  it  often  is  necessary  either  to  pit  the  stoker  or  oil  burner,  or 

408 


CHAPTER  25 — BOILERS 


where  water  line  conditions  and  headroom  permit,  to  raise  the  boiler  on  a 
brick  foundation  setting. 

Smokeless  combustion  of  the  more  volatile  bituminous  coals  is  furthered 
by  the  use  of  mechanical  stokers.  (See  Chapter  28.)  Smokeless  com- 
bustion in  hand-fired  boilers  burning  high  volatile  solid  fuel  is  aided  (1) 
by  the  use  of  double  grates  with  down-draft  through  the  upper  grate,  (2) 
by  the  use  of  a  curtain  section  through  which  preheated  auxiliary  air  is 
introduced  over  the  fire  toward  the  rear  of  the  boiler,  and  (3)  by  the  intro- 
duction of  preheated  air  through  passages  at  the  front  of  the  boiler.  All 
three  methods  depend  largely  on  mixing  secondary  air  with  the  partially 
burned  volatiles  and  causing  this  mixture  to  pass  over  an  incandescent 
fuel  bed,  thus  tending  to  secure  more  complete  combustion  than  is  pos- 
sible in  boilers  without  such  provision. 

HEATING  SURFACE 

Boiler  heating  surface  is  that  portion  of  the  surface  of  the  heat  transfer 
apparatus  in  contact  with  the  fluid  being  heated  on  one  side  and  the  gas  or 
refractory  being  cooled  on  the  other  side.  Heating  surface  on  which  the 
fire  shines  is  known  as  direct  or  radiant  surface  and  that  in  contact  with 
hot  gases  only,  as  indirect  or  convection  surface.  The  amount  of  heating 
surface,  its  distribution  and  the  temperatures  on  either  side  thereof 
influence  the  capacity  of  any  boiler. 

Direct  heating  surface  is  more  valuable  than  indirect  per  square  foot 
because  it  is  subjected  to  a  higher  temperature  and  also,  in  the  case  of 
solid  fuel,  because  it  is  in  position  to  receive  the  full  radiant  energy  of  the 
fuel  bed.  The  heat  transfer  capacity  of  a  radiant  heating  surface  may  be 
as  high  as  6  to  8  times  that  of  an  indirect  surface.  This  is  one  of  the 
reasons  why  the  water  legs  of  some  boilers  have  been  extended,  especially 
in  the  case  of  stoker  firing  where  the  extra  amount  of  combustion  chamber 
secured  by  an  extension  of  the  water  legs  is  important.  For  the  same 
reason,  care  should  be  exercised  in  building  a  refractory  combustion 
chamber  in  an  oil-burning  boiler  so  as  not  to  screen  any  more  of  this 
valuable  surface  with  refractories  than  is  necessary  for  good  combustion. 

The  effectiveness  of  the  heating  surface  depends  on  its  cleanliness,  its 
location  in  the  boiler,  and  the  shape  of  the  gas  passages.  Investigations1 
by  the  U.  S.  Bureau  of  Mines  show  that: 

1.  A  boiler  in  which  the  heating  surface  is  arranged  to  give  long  gas  passages  of  small 
cross-section  will  be  more  efficient  than  a  boiler  in  which  the  gas  passages  are  short  and  of 
larger  cross-section. 

2.  The  efficiency  of  a  water  tube  boiler  increases  as  the  free  area  between  individual 
tubes  decreases  and  as  the  length  of  the  gas  pass  increases. 

3.  By  inserting  baffles  so  that  the  heating  surface  is  arranged  in  series  with  respect  to 
the  gas  flow,  the  boiler  efficiency  will  be  increased. 

The  area  of  the  gas  passages  must  not  be  so  small  as  to  cause  excessive 
resistance  to  the  flow  of  gases  where  natural  draft  is  employed. 

Heat  Transfer  Rates 

Practical  rates  of  heat  transfer  in  heating  boilers  will  average  about 

!See  U.  S.  Bureau  of  Mines  Bulletin  No.  IS,  The  Transmission  of  Heat  into  Steam  Boilers. 

409 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

3300  Btu  per  sq  ft  per  hour  for  hand-fired  boilers  and  4000  Btu  per  sq  ft 
per  hour  for  mechanically  fired  boilers  when  operating  at  design  load2. 
When  operating  at  maximum  load2  these  values  will  run  between  5000  and 
6000  Btu  per  sq  ft  per  hour.  Boilers  operating  under  favorable  conditions 
at  the  above  heat  transfer  rates  will  give  exit  gas  temperatures  that  are 
considered  consistent  with  good  practice. 

TESTING  AND  RATING  CODES 

The  Society  has  adopted  three  solid  fuel  testing  codes,  a  solid  fuel 
rating  code  and  an  oil  fuel  testing  code.  A.S.H.V.E.  Standard  and  Short 
Form  Heat  Balance  Codes  for  Testing  Low-Pressure  Steam  Heating 
Solid  Fuel  Boilers— Codes  1  and  2 — (Revision  of  June  1929)3,  are  intended 
to  provide  a  method  for  conducting  and  reporting  tests  to  determine  heat 
efficiency  and  performance  characteristics.  A.S.H.V.E.  Performance 
Test  Code  for  Steam  Heating  Solid  Fuel  Boilers — Code  No.  3 — (Edition  of 
1929)3  is  intended  for  use  with  A.S.H.V.E.  Code  for  Rating  Steam  Heating 
Solid  Fuel  Hand-Fired  Boilers4.  The  object  of  this  test  code  is  to  specify 
the  tests  to  be  conducted  and  to  provide  a  method  for  conducting  and 
reporting  tests  to  determine  the  efficiencies  and  performance  of  the  boiler. 
The  A.S.H.V.E.  Standard  Code  for  Testing  Steam  Heating  Boilers 
Burning  Oil  Fuel5  is  intended  to  provide  a  standard  method  for  con- 
ducting and  reporting  tests  to  determine  the  heating  efficiency  and  per- 
formance characteristics  when  oil  fuel  is  used  with  steam  heating  boilers. 
The  Steel  Heating  Boiler  Institute  suggests  a  single  number  dimensional 
rating  in  the  S.H.B.I.  Code  for  the  Rating  of  Low-Pressure  Heating 
Boilers  by  Their  Physical  Characteristics6. 

BOILER  OUTPUT 

Boiler  output  as  defined  in  A.S.H.V.E.  Performance  Test  Code  for 
Steam  Heating  Solid  Fuel  Boilers  (Code  No.  3)  is  the  quantity  of  heat 
available  at  the  boiler  nozzle  with  the  boiler  normally  insulated.  It 
should  be  based  on  actual  tests  conducted  in  accordance  with  this  code. 
This  output  is  usually  stated  in  Btu  and  in  square  feet  of  equivalent  heat- 
ing surface  (radiation).  According  to  the  A.S.H.V.E.  Standard  Code  for 
Rating  Steam  Heating  Solid  Fuel  Hand-Fired  Boilers,  the  performance 
data  should  be  given  in  tabular  or  curve  form  on  the  following  items  for  at 
least  five  outputs  ranging  from  maximum  down  to  35  per  cent  of  maxi- 
mum: (1)  fuel  available,  (2)  combustion  rate,  (3)  efficiency,  (4)  draft 
tension,  (5)  flue  gas  temperature.  The  only  definite  restriction  placed  on 
setting  the  maximum  output  is  that  priming  shall  not  exceed  2  per  cent. 
These  curves  provide  complete  data  regarding  the  performance  of  the 
boiler  under  test  conditions.  Certain  other  pertinent  information,  such  as 
grate  area,  heating  surface  and  chimney  dimensions  is  desirable  also  in 
forming  an  opinion  of  how  the  boiler  will  perform  in  actual  service. 

The  output  of  large  heating  boilers  is  frequently  stated  in  terms  of 


•For  definitions  of  design  load  and  maximum  load  see  pages  411  and  412. 
•See  A.S.H.V.E.  TRANSACTIONS,  Vol.  35,  1929.    Also  Chapter  41, 
<See  A.S.H.V.E.  TRANSACTIONS,  Vol.  36, 1930.    Also  Chapter  41. 
*See  A.S.H.V.E.  TRANSACTIONS,  Vol.  37,  1931.    Also  Chapter  41. 

•See  Rating  of  Heating  Boilers  by  Their  Physical  Characteristics,  by  C.  E.  Branson  (A.S.H.V.E.  TRANS- 
ACTIONS, Vol.  36,  1930). 

410 


CHAPTER  25 — BOILERS 


boiler  horsepower  instead  of  in  Btu  per  hour  or  square  feet  of  equivalent 
radiation. 

Boiler  Horsepower:  The  evaporation  of  34.5  Ib  of  water  per  hour 
from  and  at  212  F  which  is  equivalent  to  a  heat  output  of  970.2  X  34.5  = 
33,471.9  Btu  per  hour. 

Equivalent  Evaporation:  The  amount  of  water  a  boiler  would 
evaporate,  in  pounds  per  hour,  if  it  received  feed  water  at  212  F  and 
vaporized  it  at  this  same  temperature  and  at  atmospheric  pressure. 

It  is  usually  considered  that  10  sq  ft  of  boiler  heating  surface  will  pro- 
duce a  rated  boiler  horsepower.  A  rated  boiler  horsepower  in  turn 
can  carry  a  design  load  of  from  100  to  140  sq  ft  of  equivalent  radiation. 
It  is  apparent,  therefore,  that  1  sq  ft  of  boiler  heating  surface  can  carry  a 
design  load  of  from  10  to  14  sq  ft  of  equivalent  radiation,  or  somewhat 
more  if  the  boiler  is  forced  above  rating.  The  application  of  these  values 
is  discussed  under  the  heading  Selection  of  Boilers. 

BOILER  EFFICIENCY 

The  term  efficiency  as  used  for  guarantees  of  boiler  performance  is 
usually  construed  as  follows: 

1.  Solid  Fuels.    The  efficiency  of  the  boiler  alone  is  the  ratio  of  the  heat  absorbed  by 
the  water  and  steam  in  the  boiler  per  pound  of  combustible  burned  on  the  grate  to  the 
calorific  value  of  1  Ib  of  combustible  as  fired.    The  combined,  efficiency  of  'boiler,  furnace 
and  grate  is  the  ratio  of  the  heat  absorbed  by  the  water  and  steam  in  the  boiler  per  pound 
of  fuel  as  fired  to  the  calorific  value  of  1  Ib  of  fuel  as  fired. 

2.  Liquid  Fuels.    The  combined  efficiency  of.  boiler,  furnace  and  burner  is  the  ratio  of 
the  heat  absorbed  by  the  water  and  steam  in  the  boiler  per  pound  of  fuel  to  the  calorific 
value  of  1  Ib  of  fuel. 

Solid  fuel  boilers  usually  show  an  efficiency  of  50  to  75  per  cent  when 
operated  under  favorable  conditions  at  their  rated  capacities.  Infor- 
mation on  the  combined  efficiencies  of  boiler,  furnace  and  burner  has 
resulted  from  research  conducted  at  Yale  University  in  cooperation  with 
the  A.S.H.V.E.  Research  Laboratory  and  the  American  Oil  Burner 
Association1.  For  general  information  on  heating  efficiencies  see  Chapter 
29. 

SELECTION  OF  BOILERS 

Estimated  Design  Load:  The  load,  stated  in  Btu  per  hour  or  equiv- 
alent direct  radiation,  as  estimated  by  the  purchaser  for  the  conditions  of 
inside  and  outside  temperature  for  which  the  amount  of  installed  radiation 
was  determined  is  the  sum  of  the  heat  emission  of  the  radiation  to,  be 
actually  installed  plus  the  allowance  for  the  heat  loss  of  the  connecting 
piping  plus  the  heat  requirement  for  any  apparatus  requiring  heat  con- 
nected with  the  system  (A.S.H.V.E.  Standard  Code  for  Rating  Steam 
Heating  Solid  Fuel  Hand-Fired  Boilers— Edition  of  April,  1932). 

The  estimated  design  load  is  the  sum  of  the  following  three  items8: 

1.  The  estimated  heat  emission  in  Btu  per  hour  of  the  connected  radiation  (direct, 
indirect  or  central  fan)  to  be  installed. 


'See  A.S.H.V.E.  research  papers  entitled  Study  of  the  Characteristics  of  Oil  Burners  and  Heating 
Boilers,  by  L.  E.  Seeley  and  E.  J.  Tavanlar  (A.S.H.V.E.  TRANSACTIONS,  Vol.  37,  1931),  and  A  Study  of 
Intermittent  Operation  of  Oil  Burners,  by  L.  E.  Seeley  and  J.  H.  Powers  (A.S.H.V.E.  TRANSACTIONS, 
Vol.  38,  1932). 

•See  A.S.H.V.E.  Code  of  Minimum  Requirements  for  the  Heating  and  Ventilation  of  Buildings  (Edition 
of  1929). 

411 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


2.  The  estimated  maximum  heat  in  Btu  per  hour  required  to  supply  water  heaters 
or  other  apparatus  to  be  connected  to  the  boiler. 

3.  The  estimated  heat  emission  in  Btu  per  hour  of  the  piping  connecting  the  radiation 
and  other  apparatus  to  the  boiler. 

Estimated  Maximum  Load:  Construed  to  mean  the  load  stated  in 
Btu  per  hour  or  the  equivalent  direct  radiation  that  has  been  estimated  by 
the  purchaser  to  be  the  greatest  or  maximum  load  that  the  boiler  will  be 
called  upon  to  carry.  (A.S.H.V.E.  Standard  Code  for  Rating  Steam 
Heating  Solid  Fuel  Hand-Fired  Boilers— Edition  of  April,  1932.) 

The  estimated  maximum  load  is  given  by8: 

4.  The  estimated  increase  in  the  normal  load  in  Btu  per  hour  due  to  starting  up  cold 
radiation.    This  percentage  of  increase  is  to  be  based  on  the  sum  of  Items  1,  2  and  3 
and  the  heating-up  factors  given  in  Table  2. 

TABLE  2.    WARMING-UP  ALLOWANCES  FOR  Low  PRESSURE  STEAM  AND 
HOT  WATER  HEATING  BOILERSSI  t>,  c 


DESIGN  LOAD  (REPRESENTING  SUMMATION  OF  ITEMS  1,  2,  AND  3,d 

PERCENTAGE  CAPACITT  TO  ADD 
FOR  WARMING  UP 

Btu  per  Hour 

Equivalent  Square  Feet  of  Radiationd 

.  Up  to  100,000 
100,000  to  200,000 
200,000  to  600,000 
600,000  to  1,200,000 
1,200,000  to  1,800,000 
Above  1,800,000 

Up  to  420 
420  to  840 
840  to  2500 
2500  to  5000 
5000  to  7500 
Above  7500 

65 
60 
55 
50 
45 
40 

aThis  table  is  taken  from  the  A.S.H.V.E.  Code  of  Minimum  Requirements  for  the  Heating  and  Venti- 
lation of  Buildings,  except  that  the  second  column 'has  been  added  for  convenience  in  interpreting  the  design 
load  in  terms  of  equivalent  square  feet  of  radiation. 

bSee  also  Time  Analysis  in  Starting  Heating  Apparatus,  by  Ralph  C.  Taggert  (A.S.H.V.E.  TRANSAC- 
TIONS, Vol.  19,  1913);  Report  of  A.S.H.V.E.  Continuing  Committee  on  Codes  for  Testing  and  Rating  Steam 
Heating  Solid  Fuel  Boilers  (A.S.H.V.E.  TRANSACTIONS,  yol.  36,  1930);  Selecting  the  Right  Size  Heating 
Boiler,  by  Sabin  Crocker  (Heating,  Piping  and  Air  Conditioning,  March,  1932). 

cThis  table  refers  to  hand-fired  solid  fuel  boilers.  A  factor  of  25  per  cent  over  design  load  is  adequate 
when  oil  or  gas  are  used  as  fuels. 

d240  Btu  per  square  foot. 

Other  things  to  be  considered  are: 

5.  Efficiency  with  hard  or  soft  coal,  gas,  or  oil  firing,  as  the  case  may  be. 

6.  Grate  area  with  hand-fired  coal,  or  fuel  burning  rate  with  stokers,  oil,  or  gas. 

7.  Combustion  space  in  the  furnace. 

8.  Type  of  heat  liberation,  whether  continuous  or  intermittent,  or  a  combination  of 
both. 

9.  Miscellaneous  items  consisting  of  draft  available,  character  of  attendance,  pos- 
sibility of  future  extension,  possibility  of  breakdown,  headroom  in  the  boiler  room. 

Radiation  Load 

The  connected  radiation  (Item  1)  is  determined  by  calculating  the  heat 
losses  in  accordance  with  data  given  in  Chapters  5,  6  and  7,  and  dividing 
by  240  to  change  to  square  feet  of  equivalent  radiation  as  explained  in 
Chapter  30.  For  hot  water,  the  emission  commonly  used  is  150  Btu  per 
square  foot,  but  the  actual  emission  depends  on  the  temperature  of  the 
medium  in  the  heating  units  and  of  the  surrounding  air.  (See  Chapter  30.) 

Although  it  is  customary  to  use  the  actual  connected  load  in  equivalent 
square  feet  of  radiation  for  selecting  the  size  of  boiler,  this  connected  load 
usually  represents  a  reserve  in  heating  capacity  to  provide  for  infiltration 
in  the  various  spaces  of  the  building  to  be  heated,  which  reserve,  however, 

412 


CHAPTER  25 — BOILERS 


is  not  in  use  at  all  places  at  the  same  time,  or  in  any  one  place  at  all  times. 
For  a  further  discussion  of  this  subject  see  Chapter  6. 

Hot  Water  Supply  Load 

When  the  hot  water  supply  (Item  2)  is  heated  by  the  building  heating 
boiler,  this  load  must  be  taken  into  consideration  in  sizing  the  boiler.  The 


§u. 
0       w 


iboo     2000     3000     4000     5000     eooo     7000 u 

OUTPUT,  SQ.  FT.  EQUIVALENT  STEAM  RADIATION   (240  BTU   PER  SQ.FT.) 
BOILER    DATA—      GRATE  ARE A,SQ. FT.   18-0  FUEL— BITUMINOUS    3/4"  LUMP 

HEATING  SURFACE.SQ. FT.  254  ANALYSES-   VOLATILE  MATTER       34-06% 


WEIGHT,  LB. 
FUEL  CAPACITY,  LB. 
FUEL  AVAILABLE,  LB. 
FUEL   DEPTH,  IN 


9160 

651 

414. 

10 


FIXED  CARBON 
ASH 

SULPHUR 
MOISTURE 
BTU  PERLB. 


55.44 
0.67 
2.66 
3.00 
13,655 


FIG.  1.    TYPICAL  PERFORMANCE  CURVES  FOR  A  36-iN.  CAST-IRON  SECTIONAL  STEAM 

HEATING  BOILER,  BASED  ON  THE  A.S.H.V.E.  CODE  FOR  RATING  STEAM 

HEATING  SOLID  FUEL  HAND-FIRED  BOILERS 

allowance  to  be  made  will  depend  on  the  amount  of  water  heated  and  its 
temperature  rise.  A  good  approximation  is  to  add  4  sq  ft  of  equivalent 
radiation  for  each  gallon  of  water  heated  per  hour  through  a  temperature 
range  of  100  F.  For  more  specific  information,  see  Chapter  35. 

Piping  Tax  (Item  3) 

It  is  common  practice  to  add  a  flat  percentage  allowance  to  the 
equivalent  connected  radiation  to  provide  for  the  heat  loss  from  bare  and 
covered  pipe  in  the  supply  and  return  lines.  The  use  of  a  flat  allowance  of 
25  per  cent  for  steam  systems  and  35  per  cent  for  hot  water  systems  is 
preferable  to  ignoring  entirely  the  load  due  to  heat  loss  from  the  supply 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

and  return  lines,  but  better  practice,  especially  when  there  is  much  bare 
pipe,  is  to  compute  the  emission  from  both  bare  and  covered  pipe  surface 
in  accordance  with  data  in  Chapter  36.  With  direct  radiation  served  by 
bare  supply  and  return  piping  the  percentages  may  be  higher  than  those 
stated,  while  in  the  case  of  unit  heaters  where  the  output  is  concentrated 
in  a  few  locations,  the  piping  tax  may  be  10  per  cent  or  less. 

Warming-up  Allowance 

The  warming-up  allowance  represents  the  load  due  to  heating  the  boiler 
and  contents  to  operating  temperature  and  heating  up  cold  radiation  and 
piping.  (See  Item  4.)  The  factors  to  be  used  for  determining  the 
allowance  to  be  made  should  be  selected  from  Table  2  and  should  be 
applied  to  the  estimated  design  load  as  determined  by  Items  1,  2  and  3. 

Performance  Curves  for  Boiler  Selection 

In  the  selection  of  a  boiler  to  meet  the  estimated  load,  the  A.S.H.V.E. 
Standard  Code  for  Rating  Steam  Heating  Solid  Fuel  Hand-Fired  Boilers 
recommends  the  use  of  performance  curves  based  on  actual  tests  con- 
ducted in  accordance  with  the  A.S.H.V.E.  Performance  Test  Code  for 
Steam  Heating  Solid  Fuel  Boilers  (Code  No.  3),  similar  to  the  typical 
curves  shown  in  Fig.  1.  It  should  be  understood  that  performance  data 
apply  to  test  conditions  and  that  a  reasonable  allowance  should  be  made 
for  decreased  output  resulting  from  soot  deposit,  poor  fuel  or  inefficient 
attention. 

Selection  Based  on  Heating  Surface  and  Grate  Area 

Where  performance  curves  are  not  available,  a  good  general  rule  for 
conventionally-designed  boilers  is  to  provide  1  sq  ft  of  boiler  heating 
surface  for  each  14  sq  ft  of  equivalent  radiation  (240  Btu  per  square  foot) 
represented  by  the  design  load  consisting  of  connected  radiation,  piping 
tax  and  domestic  water  heating  load.  As  stated  in  the  section  on  Boiler 
Output,  this  is  equivalent  to  allowing  10  sq  ft  of  boiler  heating  surface  per 
boiler  horsepower.  In  this  case  it  is  assumed  that  the  maximum  load 
including  the  warming-up  allowance  will  be  provided  for  by  operating  the 
boiler  in  excess  of  the  design  load,  that  is,  in  excess  of  the  100  per  cent 
rating  on  a  boiler-horsepower  basis. 

Due  to  the  wide  variation  encountered  in  manufacturers'  ratings  for 
boilers  of  approximately  the  same  capacity,  it  is  advisable  to  check  the 
grate  area  required  for  heating  boilers  burning  solid  fuel  by  means  of  the 
following  formula: 

G  =  ~CX  FXE  (1) 

where 

G  =a  grate  area,  square  feet. 

H  ~  required  total  heat  output  of  the  boiler,  Btu  per  hour  (see  Selection  of  Boilers, 

p.  411), 
C  =  combustion  rate  in  pounds  of  dry  coal  per  square  foot  of  grate  area  per  hour, 

depending  on  the  kind  of  fuel  and  size  of  boiler  as  given  in  Table  1. 
F  =  calorific  value  of  fuel,  Btu  per  pound. 
E  s=  efficiency  of  boiler,  usually  taken  as  0.60. 

Example  1.    Determine  the  grate  area  for  a  required  heat  output  of  the  boiler  of 

414 


CHAPTER  25 — BOILERS 


500,000  Btu  per  hour,  a  combustion  rate  of  6  Ib  per  hour,  a  calorific  value  of  13,000  JBtu 
per  pound,  and  an  efficiency  of  60  per  cent. 

r  _  500,000  '      1A«       r 

G  "  ex  13,000x0.60  "  10-7  sq  ft  ; ".,;;.; 

The  boiler  selected  should  have  a  grate  area  not  less  than  that  deter- 
mined by  Formula  1.  With  small  boilers  where  it  is  desired  to-  provide 
sufficient  coal  capacity  for  approximately  an  eight-hour  firing  period  plus 
a  20  per  cent  reserve  for  igniting  a  new  charge,  more  grate  area  may  be 
required  depending  upon  the  depth  of  the  fuel  pot. 

Selection  of  Gas-Fired  Boilers 

Gas-heating  appliances  should  be  selected  in  accordance  with  factors 
given  in  Table  1,  Chapter  28,  which  include  an  allowance  for  heating  up 
cold  radiation,  and  for  the  piping  tax.  These  factors  are  for  thermo-- 
statically-controlled  systems;  in  case  manual  operation  is  desired,  a 
warming-up  allowance  of  100  per  cent  is  recommended  by  the  A^G'.A: 
A  gas  boiler  selected  by  the  use  of  the  A  .G.A.  factors  will  be  the  minimum 
size  boiler  which  can  carry  the  load.  From  a  fuel  economy  standpoint,  it 
may  be  advisable  to  select  a  somewhat  larger  boiler  and  then  throttle  the 
gas  and  air  adjustments  as  required.  "This  will  tend  to'  give  a  low  stack 
temperature  with  high  efficiency  and  at  the  same  time  provide  reserve 
capacity  in  case  the  load  is  underestimated  or  more  is  added  in  the  future: 

Conversions 

In  the  case  of  a  solid  fuel  boiler  converted  to  gas  burning,  the  heat  units 
supplied  in  the  gas  should  be  approximately  twice  the 'connected  heating 
load.  A  combustion  efficiency  of  75  per  cent  for  a  conversion"  installation 
would  provide  a  boiler  output  of  2  X  0.75  =  1.5  times  the  connected  Io4d-, 
which  allows  50  per  cent  for  piping  tax  and  pick-up. .  The  presumption  for 
a  conversion"  job  is  that  the  boiler  already  is  installed  and  probably  will 
not  be  made  larger;  therefore,  it  is  a  matter  of  setting  a  gas-burning  rate 
to  obtain  best  results  with  the  available  surface.  The  conversion  of  a  coal 
or  oil  boiler  to  gas  burning  is  accomplished  much  more  rapidly  than  the 
reverse  since  but  little  furnace  volume  need  be  provided  for  the  proper 
combustion  of  gas. 

Other  Considerations  in  Selection  of  Boilers 

As  it  will  usually  be  found  that  several  boilers  will  meet  the  speci- 
fications, the  final  selection  of  the  boiler  may  be  influenced  by  other  con- 
siderations, some  of  which  are : 

1.  Dimensions  of  boiler.  .  .  . 

2.  Durability  under  service. 

3.  Convenience  in  'firing  and  cleaning.  '   .    '     "     -    *  -  -    - 

4.  Adaptability  to  changes  in  fuel  and  kind  of  attention. 

5.  Height  of  water  line.  .  :  \ '  ,     .  ' .  '    ' ' 

In  large  installations,  the  use  of  several  smaller  boiler  units  instead  of 
one  larger  one  will  obtain  greater  flexibility  and  economy  by, permitting 
the  operation,  at  the  best  efficiency,  of  the  required  number  of  units 
according  to  the  heat  requirements.  •  -  •  _ 

Boiler  rooms  should,  if  possible,  be  situated  at  a  central  point  with 

415 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

respect  to  the  building  and  should  be  designed  for  a  maximum  of  natural 
light.  The  space  in  front  of  the  boilers  should  be  sufficient  for  firing, 
stoking,  ash  removal  and  cleaning  or  renewal  of  flues,  and  should  be  at 
least  3  ft  greater  than  the  length  of  the  boiler  firebox. 

A  space  of  at  least  3  ft  should  be  allowed  on  at  least  one  side  of  every 
boiler  for  convenience  of  erection  and  for  accessibility  to  the  various 
dampers,  cleanouts  and  trimmings.  The  space  at  the  rear  of  the  boiler 
should  be  ample  for  the  chimney  connection  and  for  cleanouts,  and  with 
large  boilers  the  rear  clearance  should  be  at  least  3  ft  in  width. 

The  boiler  room  height  should  be  sufficient  for  the  location  of  boiler 
accessories  and  for  proper  installation  of  piping.  In  general  the  ceiling 
height  for  small  steam  boilers  should  be  at  least  3  ft  above  the  normal 
boiler  water  line.  With  vapor  heating,  especially,  the  height  above  the 
boiler  water  line  is  of  vital  importance. 

When  steel  boilers  are  used,  space  should  be  provided  for  the  removal 
and  replacement  of  tubes. 

CONNECTIONS  AND  FITTINGS 

The  velocity  of  flow  through  the  outlets  of  low  pressure  steam  heating 
boilers  should  not  exceed  15  to  25  fps  if  fluctuation  of  the  water  line  and 
undue  en  trainmen  t  of  moisture  are  to  be  avoided.  Steam  or  water  outlet 
connections  preferably  should  be  the  full  size  of  the  manufacturers' 
tapping  and  should  extend  vertically  to  the  maximum  height  available 
above  the  boiler.  For  gravity  circulating  steam  heating  systems,  it  is 
recommended  that  a  Hartford  Loop,  described  in  Chapter  32,  be  utilized 
in  making  the  return  connection. 

Particular  attention  should  be  given  to  fitting  connections  to  secure  con- 
formity with  the  A.S.M.E.  Boiler  Construction  Code  for  Low  Pressure 
Heating  Boilers.  Attention  is  called  in  particular  to  pressure  gage  piping, 
water  gage  connections  and  safety  valve  capacity. 

Steam  gages  should  be  fitted  with  a  water  seal  and  a  shut-off  consisting 
of  a  cock  with  either  a  tee  or  lever  handle  which  is  parallel  to  the  pipe 
when  the  cock  is  open.  Steam  'gage  connections  should  be  of  copper  or 
brass  when  smaller  than  1  in.  J.P.5.9  if  the  gage  is  more  than  5  ft  from  the 
boiler  connection,  and  also  in  any  case  where  the  connection  is  less  than 
%  in.  I.P.S. 

Each  steam  or  vapor  boiler  should  have  at  least  one  water  gage  glass  and 
two  or  more  gage  cocks  located  within  the  range  of  the  visible  length  of  the 
glass.  The  water  gage  fittings  or  gage  cocks  may  be  direct  connected  to 
the  boiler,  if  so  located  by  the  manufacturer,  or  may  be  mounted  on  a 
separate  water  column.  No  connections,  except  for  combustion  regu- 
lators, drains  or  steam  gages,  should  be  placed  on  the  pipes  connecting 
the  water  column  and  the  boiler.  If  the  water  column  or  gage  glass  is  con- 
nected to  the  boiler  by  pipe  and  fittings,  a  cross,  tee  or  equivalent,  in  which 
a  cleanout  plug  or  a  drain  valve  and  piping  may  be  attached,  should  be 
placed  in  the  water  connection  at  every  right-angle  turn  to  facilitate 
cleaning.  The  water  line  in  steam  boilers  should  be  carried  at  the  level 
specified  by  the  boiler  manufacturer. 


.  Code,  Identification  of  Piping  Systems. 

416 


CHAPTER  25 — BOILERS 


Safety  valves  should  be  capable  of  discharging  all  the  steam  that  can  be 
generated  by  the  boiler  without  allowing  the  pressure  to  rise  more  than 
5  Ib  above  the  maximum  allowable  working  pressure  of  the  boiler.  This 
should  be  borne  in  mind  particularly  in  the  case  of  boilers  equipped  with 
mechanical  stokers  or  oil  burners  where  the  amount  of  grate  area  has 
little  significance  as  to  the  steam  generating  capacity  of  the  boiler. 

Where  a  return  header  is  used  on  a  cast-iron  sectional  boiler  to  distribute 
the  returns  to  both  rear  tappings,  it  is  advisable  to  provide  full  size 
plugged  tees  instead  of  elbows  where  the  branch  connections  enter  the 
return  tappings.  This  facilitates  cleaning  sludge  from  the  bottom  of  the 
boiler  sections  through  the  large  plugged  openings.  An  equivalent  clean- 
out  plug  should  be  provided  in  the  case  of  a  single  return  connection. 

Blow-off  or  drain  connections  should  be  made  near  the  boiler  and  so 
arranged  that  the  entire  system  may  be  drained  of  water  by  opening  the 
drain  cock.  In  the  case  of  two  or  more  boilers  separate  blow-off  connec- 
tions must  be  provided  for  each  boiler  on  the  boiler  side  of  the  stop  valve 
on  the  main  return  connection. 

Water  service  connections  must  be  provided  for  both  steam  and  water 
boilers,  for  refilling  and  for  the  addition  of  make-up  water  to  boilers.  This 
connection  is  usually  of  galvanized  steel  pipe,  and  is  made  to  the  return 
main  near  the  boiler  or  boilers. 

For  further  data  on  pipe  connections  for  steam  and  hot  water  heating 
systems,  see  Chapters  32  and  33  and  the  A.S.M.E.  Boiler  Construction 
Code  for  Low  Pressure  Heating  Boilers. 

Smoke  Breeching  and  Chimney  Connections.  The  breeching  or  smoke 
pipe  from  the  boiler  outlet  to  the  chimney  should  be  air-tight  and  as  short 
and  direct  as  possible,  preference  being  given  to  long  radius  and  45-deg 
instead  of  90-deg  bends.  The  breeching  entering  a  brick  chimney  should 
not  project  beyond  the  flue  lining  and  where  practicable  it  should  be 
grouted  up  from  the  inside  of  the  chimney.  A  thimble  or  sleeve  grout 
usually  is  provided  where  the  breeching  enters  a  brick  chimney. 

Where  a  battery  of  boilers  is  connected  into  a  breeching  each  boiler 
should  be  provided  with  a  tight  damper.  The  breeching  for  a  battery 
of  boilers  should  not  be  reduced  in  size  as  it  goes  to  the  more  remote 
boilers.  Good  connections  made  to  a  good  chimney  will  usually  result  in 
a  rapid  response  by  the  boilers  to  demands  for  heat. 

ERECTION,  OPERATION,  AND  MAINTENANCE 

The  directions  of  the  boiler  manufacturer  always  should  be  read  before 
the  assembly  or  installation  of  any  boiler  is  started,  even  though  the 
contractor  may  be  familiar  with  the  boiler.  All  joints  requiring  boiler 
putty  or  cement  which  cannot  be  reached  after  assembly  is  complete 
must  be  finished  as  the  assembly  progresses. 

The  following  precautions  should  be  taken  in  all  installations  to  prevent 
damage  to  the  boiler: 

1.  There  should  be  provided  proper  and  convenient  drainage  connections  for  use  if 
the  boiler  is  not  in  operation  during  freezing  weather. 

2.  Strains  on  the  boiler  due  to  movement  of  piping  during  expansion  should  be 
prevented  by  suitable  anchoring  of  piping  and  by  proper  provision  for  pipe  expansion 
and  contraction. 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

3.  Direct  impingement  of  too  intense  local  heat  upon  any  part  of  the  boiler  surface, 
as  with  .oil  burners,  should  be  avoided  by  protecting  the  surface  with  firebrick  or  other 
refractory  material.  - 

4.  Condensation  must  flow  back  to  the  boiler  as  rapidly  and  uniformly  as  possible. 
Return  connections  should  prevent  the  water  from  backing  out  of  the  boiler. 

5.  Automatic  boiler  feeders  and  low  water  cut-off  devices  which  shut  off  the  source 
of  heat  if  the  water  in  the  boiler  falls  below  a  safe  level  are  recommended  for  boilers 
mechanically  fired. 

Boiler  Troubles 

A  complaint  regarding  boiler  operation  generally  will  be  found  to  be 
due  to  one  of  the  following: 

-  1.  The  boiler  fails  to  deliver  enough  heat.  The  cause  of  this  condition  may  be:  (a)  poor 
draft -r  (b)  poor  fuel;  (c}  inferior  attention  or  firing;  (d)  boiler  too  small;  (e)  improper 
piping;  (/)  improper  arrangement  of  sections;  (g)  heating  surfaces  covered  with  soot; 
and  (h)  insufficient  radiation  installed. 

2.  The  water  line  is  unsteady.  The  cause  of  this  condition  may  be :  (a)  grease  and  dirt 
in  boiler;  (&)  water  column  connected  to  a  very  active  section  and,  therefore,  not 
showing  actual  water  level  in  boiler;  (c)  boiler  operating  at  excessive  output. 
.  3.  Water  disappears  from  gage  glass.  This  may  be  caused  by:  (a)  priming  due  to 
grease  and  dirt  in  boiler;  (&)  too  great  pressure  difference  between  supply  and  return 
piping  preventing  return  of  condensation;  (c)  valve  closed  in  return  line;  (d)  connection 
of  bottom  of  water  column  into  a  very  active  section  or  thin  waterway;  (e)  improper 
connections  between  boilers  in  battery  permitting  boiler  with  excess  pressure  to  push 
returning  condensation  into  boiler  with  lower  pressure. 

-  4,   Water  is^carried  over  into  steam  main.   This  may  be  caused  by:    (a)  grease  and  dirt 
in  boiler;  (&)  insufficient  steam  dome  or  too  small  steam  liberating  area;  (c)  outlet  con- 
nections of  too  small  area;  (d)  excessive  rate  of  output;  (e)  water  level  carried  higher 
than  specified. 

5.  Boiler  is  slow  in  response  to  operation  of  dampers.  This  may  be  due  to :  (a)  poor 
draft  due  to  air  leaks  into  chimney  or  breeching;  (b)  inferior  fuel;  (c)  inferior  attention; 
(d)  accumulation  of  clinker  on  grate;  (e)  boiler  too  small  for  the  load. 

-  6.  Boiler  requires  too  frequent  cleaning  of  flues.    This  may  be  due  to:    (a)  poor  draft; 
(b)  smoky  combustion;  (c)  too  low  a  rate  of  combustion;  (d)  too  much  excess  air  in 
firebox  causing  chilling  of  gases. 

1.  Boiler  smokes  through  fire  door.  This  may  be  due  to:  (a)  defective  draft  in  chimney 
or  incorrect  setting  of  dampers;  (b)  air  leaks  into  boiler  or  breeching;  (c)  gas  outlet  from 
firebox  plugged  with  fuel;  (d)  dirty  or  clogged  flues;  (e}  improper  reduction  in  breeching 
size. 

Cleaning  Steam  Boilers 

All  boilers  are  provided  with  flue  clean-out  openings  through  which  the 
heating  surface  can  be  reached  by  means  of  brushes  or  scrapers.  Flues 
of  solid  fuel  boilers  should  be  cleaned  often  to  keep  the  surfaces  free  of 
soot  or  ash.  Gas  boiler  flues  and  burners  should  be  cleaned  at  least  once 
a  year.  Oil  burning  boiler  flues  should  be  examined  periodically  to  deter- 
mine when  cleaning  is  necessary. 

The  grease  used  to  lubricate  the  cutting  tools  during  erection  of  new 
piping  systems  serves  as  a  carrier  for  sand  and  dirt,  with  the  result  that 
a  scum  of  fine  particles  and  grease  accumulates  on  the  surface  of  the 
water  in -all  new  boilers,  while  heavier  particles  may  settle  to  the  bottom 
of  the  boiler  and  form  sludge.  These  impurities  have  a  tendency  to  cause 
foaming,  preventing  the  generation  of  steam  and  causing  an  unsteady 
water  line. 

This  unavoidable  accumulation  of  oil  and  grease  should  be  removed 
by  blowing  off  the  boiler  as  follows:  If  not  already  provided,  install  a 

418 


CHAPTER  25 — BOILERS 


surface  blow  connection  of  at  least  1J^  in.  nominal  pipe  size  with  outlet 
extended  to  within  18  in.  of  the  floor  or  to  sewer,  inserting  a  valve  in  line 
close  to  boiler.  Bring  the  water  line  to  center  of  outlet,  raise  steam  pres- 
sure and  while  fire  is  burning  briskly  open  valve  in  blow-off  line.  When; 
pressure  recedes  close  valve  and  repeat  process  adding  water  at  intervals 
to  maintain  proper  level.  As  a  final  operation  bring  the  pressure  in  the 
boiler  to  about  10  Ib,  close  blow-off,  draw  the  fire  or  stop  burner,  and  open 
drain  valve.  After  boiler  has  cooled  partly,  fill  and  flush  out  several  times 
before  filling  it  to  proper  water  level  for  normal  service.  The  use  of  acids, 
alkalis  and  salts  for  cleaning  is  not  favored  by  boiler  manufacturers 
because  of  the  difficulty  of  complete  removal  and  the  possibility  of  sub-, 
sequent  injury. 

Insoluble  compounds  have  been  developed  which  are  effective,  but 
special  instructions  on  the  proper  cleaning  compound  and  directions  for 
its  use  in  a  boiler,  as  given  by  the  boiler  manufacturer,  should  be  carefully 
followed. 

When  soda  ash  solution  is  to  be  used  the  procedure  is  to  add  about  5  Ib 
of  soda  ash  for  each  1000  sq  ft  of  connected  radiation.  Fill  the  boiler  with 
water  until  it  just  overflows  from  the  surface  blow  outlet  pipe  and  then 
fire  sufficiently  to  raise  the  water  temperature  to  the  boiling  point  without 
getting  up  steam  pressure.  Crack  the  boiler  feed  valve  so  that  a  steady 
trickle  will  run  out  of  the  overflow  pipe.  Allow  the  boiler  to  simmer  from 
2  to  4  hours.  At  the  end  of  this  time  the  grease  and  sediment  should  have 
passed  off  through  the  overflow  pipe  or  loosened  sufficiently  to  drain  off 
through  the  bottom  blow.  Extinguish  the  fire — preferably  by  letting  it 
burn  out  and  then  dumping  any  live  coals  into  the  ashpit  where  water  can 
be  applied  with  a  hose — and  open  the  bottom  blow  wide.  Rinse  with 
fresh  water  and  refill  to  the  normal  water  level.  If  the  water  in  the  gage 
glass  then  does  not  show  clear,  repeat  the  process  using  a  stronger  soda 
ash  solution  and  boiling  for  a  longer  time.  It  sometimes  is  necessary  to 
repeat  this  process  several  times  to  completely  rid  the  boiler  of  grease. 
Failure  to  thoroughly  eliminate  grease  usually  results  in  an  unsteady 
water  line  and  danger  of  damaging  the  boiler  through  having  the  crown- 
sheet  uncovered. 

It  is  common  practice  when  starting  new  installations  to  discharge 
heating  returns  to  the  sewer  during  the  first  week  of  operation.  This 
prevents  the  passage  of  grease,  dirt  or  other  foreign  matter  into  the  boiler 
and  consequently  may  avoid  the  necessity  of  cleaning  the  boiler.  During 
the  time  the  returns  are  being  passed  to  the  sewer,  the  feed  valve  should 
be  cracked  sufficiently  to  maintain  the  proper  water  level  in  the  boiler. 

Care  of  Idle  Heating  Boilers 

Heating  boilers  are  often  seriously  damaged  during  summer  months 
due  chiefly  to  corrosion  resulting  from  the  combination  of  sulphur  from 
the  fuel  with  the  moisture  in  the  cellar  air.  At  the  end  of  the  heating 
season  the  following  precautions  should  be  taken : 

1.  All  heating  surfaces  should  be  cleaned  thoroughly  of  soot,  ash  and  residue,  and  the 
heating  surfaces  of  steel  boilers  should  be  given  a  coating  of  lubricating  oil  on  the  fire 
side. 

2.  All  machined  surfaces  should  be  coated  with  oil  or  grease. 

419 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

3.  Connections  to  the  chimney  should  be  cleaned  and  in  case  of  small  boilers  the  pipe 
should  be  placed  in  a  dry  place  after  cleaning. 

4.  If  there  is  much  moisture  in  the  boiler  room,  it  is  desirable  to  drain  the  boiler  to 
prevent  atmospheric  condensation  on  the  heating  surfaces  of  the  boiler  when  they  are 
below  the  dew-point  temperature.    Due  to  the  hazard  of  some  one  inadvertently  building 
a  fire  in  a  dry  boiler,  however,  it  is  safer  to  keep  the  boiler  filled  with  water.    A  hot  water 
system  usually  is  left  filled  to  the  expansion  tank. 

5.  The  grates  and  ashpit  should  be  cleaned. 

6.  Clean  and  repack  the  gage  glass  if  necessary, 

7.  Remove  any  rust  or  other  deposit  from  exposed  surfaces  by  scraping  with  a  wire 
brush  or  sandpaper.    After  boiler  is  thoroughly  cleaned,  apply  a  coat  of  preservative 
paint  where  required  to  external  parts  normally  painted. 

8.  Inspect  all  accessories  of  the  boiler  carefully  to  see  that  they  are  in  good  working 
order.    In  this  connection,  oil  all  door  hinges,  damper  bearings  and  regulator  parts. 

BOILER  INSULATION 

Insulation  for  cast-iron  boilers  is  of  two  general  types:  (1)  plastic 
material  or  blocks  wired  on,  cemented  and  covered  with  canvas  or  duck; 
and  (2)  blocks,  sheets  or  plastic  material  covered  with  a  metal  jacket 
furnished  by  the  boiler  manufacturer.  Self-contained  steel  firebox  boilers 
usually  are  insulated  with  blocks,  cement  and  canvas,  or  rock  wool 
blankets;  HRT  boilers  are  brick  set  and  do  not  require  insulation  beyond 
that  provided  in  the  setting.  It  is  essential  that  the  insulation  on  a  boiler 
and  adjacent  piping  be  of  non-combustible  material  as  even  slow-burning 
insulation  constitutes  a  dangerous  fire  hazard  in  case  of  low  water  in 
the  boiler. 

REFERENCES 

A.S.H.V.E.  Code  of  Minimum  Requirements  for  the  Heating  and  Ventilation  of 
Buildings. 

A.S.H.V.E.  Standard  and  Short  Form  Heat  Balance  Codes  for  Testing  Low-Pressure 
Steam  Heating  Solid  Fuel  Boilers  (Codes  1  and  2). 

A.S.H.V.E.  Performance  Test  Code  for  Steam  Heating  Solid  Fuel  Boilers  (Code  No.  3). 

A.S.H.V.E.  Standard  Code  for  Rating  Steam  Heating  Solid  Fuel  Hand-Fired  Boilers. 

Heating,  Ventilating  and  Air  Conditioning,  by  Harding  and  Willard,  Revised  Edition, 
1932. 

A.S.M.E.  Boiler  Construction  Code  for  Low  Pressure  Heating  Boilers. 

Heating  and  Piping  Contractors  National  Association  Standards  (boiler  selection 
tables). 

House-Heating,  published  by  American  Gas  Association. 

Handbook  of  Oil  Burning,  published  by  American  Oil  Burner  Association. 

Heating  and  Ventilation,  by  Allen  and  Walker  (3rd  Edition), 

Selecting  the  Right  Size  Boiler,  by  Sab  in  Crocker  (Heating,  Piping  and  Air  Con- 
ditioning, February,  March,  April,  1932). 


PROBLEMS  IN  PRACTICE 

1  •  What  is  meant  by  a  low  pressure  heating  holler? 

A  low  pressure  heating  boiler  is  a  boiler  designed  to  be  operated  at  less  than  15  Ib  steam 
pressure  or  30  Ib  water  pressure  as  measured  by  a  gage  attached  directly  to  the  boiler. 

420 


CHAPTER  25 — BOILERS 


2  •  Name  the  construction  materials  that  distinguish  two  types  of  low  pressure 
heating  hoilers. 

a.  Cast-iron. 

b.  Steel. 

3  •  What  is  the  normal  rating  range  of  each  type  of  hoiler? 

a.  Cast-iron  boilers  are  rated  at  from  200  to  18,000  sq  ft  EDR. 

b.  Steel  boilers  are  rated  at  from  300  to  50,000  sq  ft  EDR. 

4  •  a.  What  is  meant  by  direct  boiler  heating  surface? 

b.  What  is  meant  by  indirect  boiler  heating  surface? 

a.  Direct  boiler  heating  surface  is  that  boiler  surface  upon  which  the  fire  shines,  namely, 
the  walls  of  the  firebox  and  the  crown  sheet. 

b.  Indirect  boiler  heating  surface  is  that  boiler  surface  not  exposed  to  the  direct  rays  of 
the  fire  and  over  which  heated  gases  pass  after  they  have  been  in  contact  with  the 
direct  surface.     Indirect  surface  is  generally  known  as  convective  surface. 

5  •  What  is  the  average  heat  transmission  rate  in  heating  boilers  in  Btu  per 
sq  ft  of  heating  surface  per  hour? 

3500  for  coal  burning  boilers;  4200  for  oil  burning  boilers. 

6  •  What  factors  contribute  to  economical  fuel   operation   in  low  pressure 
boilers  burning  coal  or  oil? 

a.  Proper  furnace  volume  for  complete  combustion. 

b.  Arrangement  of  heating  surfaces  in  series  to  create  a  turbulent  and  scrubbing  contact 
of  gases  against  the  convective  surfaces. 

c.  Rapid  internal  water  circulation  which  will  remove  steam  bubbles  from  the  water 
side  of  heating  surfaces  and  allow  other  steam  bubbles  to  be  formed.     Rapid  disen- 
gagement of  steam  bubbles  increases  the  steam  generating  efficiency  of  each  unit 
area  of  heating  surface,  and  thereby  lowers  flue  gas  temperatures. 

7  •  What  equipment  is  usually  directly  attached  to  a  low  pressure  heating 
hoiler? 

For  coal  burning  steam  boilers:  water  column,  water  gage,  tri-cocks,  steam  gage,  lever 
pop  safety  valve,  boiler  damper  regulator. 

For  coal  burning  hot  water  boilers:  damper  regulator,  altitude  gage,  thermometer,  relief 
valve. 

For  oil  burning  boilers,  the  damper  regulators  are  omitted  and  the  following  additional 
equipment  is  usually  attached:  automatic  water  feeder,  low  water  cutout,,  a  pressure 
control,  and  a  water  temperature  control  These  are  generally  furnished  by  the  oil 
burner  manufacturer  and  do  not  come  with  the  boiler. 


I  •  What  general  precautions  regarding  the  boiler  sho 
uire  a  proposed  heating  installation  will  work  properly? 


should  he  taken  to  make 


a.  Select  the  right  size  and  type  of  boiler. 

b.  Be  sure  the  combustion  space  is  proper  for  the  type  of  fuel  burned. 

c.  Allow  sufficient  space  around  the  boiler  for  cleaning. 

d.  Secure  proper  height  and  area  of  chimney  and  connecting  breeching. 

e.  Clean  the  boiler  thoroughly  and  provide  surface  blowoff  connections  and  bottom 
blowoff  connections  for  periodic  cleaning  after  operation  is  begun, 

/.  See  that  the  boiler  heating  surface  is  cleaned  at  regular  periods. 

g.  Check  flue  gas  temperatures  and  make  a  flue  gas  analysis  at  least  once  a  month . 

h.  Secure  information  and  advice  from  boiler  manufacturer. 

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__  AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

9~* -"Below  what  temperature ^should  the  water  in  direct  water  heaters  he  main- 
tained to  reduce  scale  formation  and  corrosion? 

140  F. 

10  •  a.  What  is  the  heat  equivalent  of  a  boiler  horsepower? 

b.  How  many  -square  feet  of  heating  surface  are  usually  required  per 
boiler  horsepower? 

a.  33,471.9  Btu  per  hour. 

b.  lOsqft. 

11  •  What  is  meant  by  equivalent  evaporation? 

The  amount  of  water  that  a  boiler  would  evaporate  per  hour,  if  the  feed  water  were  at 
212  F  and  if  the  steam  were  evaporated  at  that  temperature;  this  is  usually  spoken  of  as 
"from  and  at  212  F." 

12  •  What  loads  must  be  considered  in  determining  the  boiler  capacity  re- 
quired for  a  given  installation? 

Radiation  load. 

Hot  water  supply  load. 

Piping  tax. 

Warming-up  allowance. 

Load  allowance  for  inefficient  firing. 

13  •  A  boiler  has  6  sq  ft  of  grate  area  with  a  possible  depth  of  fuel  bed  of  18  in. 
The  fuel  burned  is  bituminous  coal  with  a  heat  value  of  12,500  Btu  per  Ib, 
The  efficiency  is  assumed  to  be  50  per  cent.    How  great  a  maximum  load  will 
this  boiler  carry  if  it  is  to  be  fired  every  8  hours  and  if  20  per  cent  of  the  fuel  is 
to.  be  left  over  to  kindle  the  next  charge? 

Volume  of  fuel  bed  =  6  X  1.5  =  9.0  cu  ft. 
Available  volume  =  0.80  X  9.0  =  7.2  cu  ft.      - 
Weight  of  available  fuel  =  40  X  7.2  =  288.0  Ib. 

288  0 
Fuel 'burned  per  hour  =  — ~-   =  36.0  Ib. 

o 

Heat  released  =  36.0  X  12,500  X  0.50  «  225,000  Btu  per  hour. 

:     •  225  000 

Maximum  load  —  — ^Tj —  ^  ^38  equivalent  square  feet. 

14  •  What  are  the  usual  causes  of  unsteady  water  line  and  priming? 

Grease  and  dirt  in  boiler. 

Overload,  .resulting  in  insufficient  steam  liberating  area. 

Small  outlet  connections. 

15  •  What  type  of  return  connection  can  be  used  for  gravity  steam  heating 
systems  to  make  'the  use  of  check  valves  unnecessary? 

The  Hartford  Loop.    (See  Chapter  32.) 


422 


Chapter  26 

CHIMNEYS  AND  DRAFT 
CALCULATIONS 

Natural    Draft,    Mechanical    Draft,     Characteristics    of    Natural 

Draft  Chimneys,  Determining  Chimney  Sizes,  General  Equation, 

Chimney  Construction,  Chimneys  for  Gas  Heating 

THE  design  and  construction  of  a  chimney  is  so  important  a  part 
of  the  heating  engineer's  work  that  a  general  knowledge  of  draft 
characterises  and  calculations  is  essential. 

Draft,  in  general,  may  be  defined  as  the  pressure  difference  between  the 
atmospheric  pressure  and  that  at  any  part  of  an  installation  through 
which  the  gases  flow.  Since  a  pressure  difference  implies  a  head,  draft 
is  a  static  force.  While  no  element  of  motion  is  inferred,  yet  motion 
in  the  form  of  circulation  of  gases  throughout  an  entire  boiler  plant 
installation  is  the  direct  result  of  draft.  This  motion  is  due  to  the  pressure 
difference,  or  unbalanced  pressure,  which  compels  the  gases  to  flow.  Draft 
is  often  classified  into  two  kinds  according  to  whether  it  is  created 
thermally  or  artificially,  viz,  (1)  natural  or  thermal  draft,  and  (2)  arti- 
ficial or  mechanical  draft. 

Natural  Draft 

Natural  draft  is  the  difference  in  pressure  produced  by  the  difference  in 
weight  between  the  relatively  hot  gases  inside  a  natural  draft  chimney  and 
an  equivalent  column  of  the  cooler  outside  air,  or  atmosphere.  Natural 
draft,  in  other  words,  is  an  unbalanced  pressure  produced  thermally  by  a 
natural  draft  chimney  as  the  pressure  transformer  and  a  temperature 
difference.  The  intensity  of  natural  draft  depends,  for  the  most  part, 
upon  the  height  of  the  chimney  above  the  grate  bar  level  and  also  the 
temperature  difference  between  the  chimney  gases  and  the  atmosphere. 

A  typical  natural  draft  system  consists  essentially  of  a  relatively  tall 
chimney  built  of  steel,  brick  or  reinforced  concrete,  operating  with  the 
relatively  hot  gases  which  have  passed  through  the  boilers  and  accessories 
and  from  which  all  of  the  heat  has  not  been  extracted.  Hot  gases  are  an 
essential  element  in  the  operation  of  a  natural  draft  system. 

A  natural  draft  chimney  performs  the  two-fold  service  of  assisting  in 
the  creation  of  draft  by  aspiration  and  also  of  discharging  the  gases  at  an 
elevation  sufficient  to  prevent  them  from  becoming  a  nuisance. 

Natural  draft  is  quite  advantageous  in  installations  where  the  total  loss 
of  draft  due  to  resistances  is  relatively  low  and  also  in  plants  which  have 
practically  a  constant  load  and  whose  boilers  are  seldom  operated  above 
their  normal  rating.  Natural  draft  systems  have  been,  and  are  still  being, 
employed  in  the  operation  of  large  plants  during  the  periods  when  the 
boilers  are  operated  only  up  to  their  normal  rating.  When  the  rate  of 

423 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


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424 


CHAPTER  26  —  CHIMNEYS  AND  DRAFT  CALCULATIONS 

operation  is  increased  above  their  normal  rating,  some  form  of  mechanical 
draft  is  employed  as  an  auxiliary  to  overcome  the  increased  resistances  or 
draft  losses.  Natural  draft  systems  are  used  almost  exclusively  in  the 
smaller  size  plants  where  the  amount  of  gases  generated  is  relatively  small 
and  it  would  be  expensive  to  install  and  operate  a  mechanical  draft 
system. 

The  principal  advantages  of  natural  draft  systems  may  be  summarized 
as  follows:  (1)  simplicity,  (2)  reliability,  (3)  freedom  from  mechanical 
parts,  (4)  low  cost  of  maintenance,  (5)  relatively  long  life,  (6)  relatively 
low  depreciation,  and  (7)  no  power  required  to  operate.  The  principal 
disadvantages  are:  (1)  lack  of  flexibility,  (2)  irregularity,  (3)  affected 
by  surroundings,  and  (4)  affected  by  temperature  changes. 

Mechanical  Draft 

Artificial  draft,  or  mechanical  draft,  as  it  is  more  commonly  called,  is  a 
difference  in  pressure  produced  either  directly  or  indirectly  by  a  forced 
draft  fan,  an  induced  draft  fan,  or  aVenturi  chimney  as  the  pressure 
transformer.  The  intensity  of  mechanical  draft  is  dependent  for  the  most 
part  upon  the  size  of  the  fan  and  the  speed  at  which  it  is  operated.  The 
element  of  temperature  does  not  enter  into  the  creation  of  mechanical 
draft  and  therefore  its  intensity,  unlike  natural  draft,  is  independent  of  the 
temperature  of  the  gases  and  the  atmosphere.  Mechanical  draft  includes 
the  induced  and  Venturi  types  of  draft  systems  in  which  the  pressure 
difference  is  the  result  of  a  suction  and  also  the  forced  draft  system  in 
which  the  pressure  difference  is  the  result  of  a  blowing.  Mechanical  draft 
systems  tend  to  produce  a  vacuum  or  a  plenum,  according  as  the  system 
used  in  its  production  creates  a  pressure  difference  below,  or  above, 
atmospheric  pressure,  respectively.  A  mechanical  draft  system  may  be 
used  either  in  conjunction  with,  or  as  an  adjunct  to,  a  natural  draft 
system. 

CHARACTERISTICS  OF  CHIMNEYS 

In  order  to  analyze  the  performance  of  a  natural  draft  chimney,  it  is 
advantageous  to  compare  its  general  operating  characteristics  with  those 
of  a  centrifugal  pump  and  also  a  centrifugally-induced  draft  fan,  there 
being  a  close  similarity  among  the  three.  Figs.  1,  2  and  3  show  the 
general  operating  characteristics  of  a  typical  centrifugally-induced  draft 
fan,  a  typical  centrifugal  pump,  and  a  typical  natural  draft  chimney, 
respectively.  The  draft-capacity  curve  of  the  chimney  corresponds  to  the 
head-capacity  curve  of  the  pump  and  also  to  the  dynamic-head  capacity  of 
the  fan  ;  the  efficiency  curve  of  the  chimney  to  the  efficiency  curves  of  the 
pump  and  fan  ;  and  the  gas  horsepower  curve  of  the  chimney  to  the  brake 
horsepower  curves  of  the  pump  and  fan. 

When  the  gases  in  the  chimney  are  stationary,  the  draft  created  is 
termed  the  theoretical  draft.  When  the  gases  are  flowing,  the  theoretical 
intensity  is  diminished  by  the  draft  loss  due  to  friction,  the  difference 
between  the  two  being  termed  the  available  draft.  The  general  equation 
for  the  available  draft  intensity  of  a  natural  draft  chimney  with  a  circular 
section  is  as  follows: 


425 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

where 

Z?a  =  available  draft,  inches  of  water. 
H  =  height  of  chimney  above  grate  bars,  feet. 
Bo  =  barometric  pressure  corresponding  to  altitude,  inches  of  mercury. 

W0  =  unit  weight  of  a  cubic  foot  of  air  at  0  F  and  sea  level  atmospheric  pressure, 
pounds  per  cubic  foot. 

We  =  unit  weight  of  a  cubic  foot  of  chimney  gases  at  0  F  and  sea  level  atmospheric 
pressure,  pounds  per  cubic  foot. 

To  «  absolute  temperature  of  atmosphere,  degrees  Fahrenheit. 
TC  =  absolute  temperature  of  chimney  gases,  degrees  Fahrenheit. 

W  =  amount  of  gases  generated  in  the  combustion  chamber  of  the  boiler  and  passing 
through  the  chimney,  pounds  per  second. 

/  ~  coefficient  of  friction. 

L  =  length  of  friction  duct  of  the  chimney,  feet. 

D  =  minimum  diameter  of  chimney,  feet. 

The  first  term  of  the  right  hand  expression  of  Equation  1  represents 
the  theoretical  draft  intensity,  and  the  second  term,  the  loss  due  to  friction. 

Example  1 .  Determine  the  available  draft  of  a  natural  draft  chimney  200  ft  in  height 
and  10  ft  in  diameter  operating  under  the  following  conditions  ^atmospheric  tempera- 
ture, 62  F ;  chimney  gas  temperature,  500  F ;  sea  level  atmospheric  pressure,  B0  =  29.92 
in.  of  mercury;  atmospheric  and  chimney  gas  density,  0.0863  and  0.09,  respectively; 
coefficient  of  friction,  0.016;  length  of  friction  duct,  200  ft.  The  chimney  discharges 
100  Ib  of  gases  per  second. 

Substituting  these  values  in  Equation  1  and  reducing: 
D    -296X200X2992xf^§       ^\       0-00126  X  100'  X  960  X  0.016  X  200 

^a  -  ^b  x  zw  x  ^y.y^  x  ^  522    -  960  j  1Q6  x  29  92  x  0  09 

=  1.27  -  0.14  =  1.13  in. 

Fig.  4  shows  the  variation  in  the  available  draft  of  a  typical  200  ft  by 
10  ft  chimney  operating  under  the  general  conditions  noted  in  Example  1. 
When  the  chimney  is  under  static  conditions  and  no  gases  are  flowing,  the 
available  draft  is  equal  to  1.27  in.  of  water,  the  theoretical  intensity.  As 
the  amount  of  gases  flowing  increases,  the  available  intensity  decreases 
until  it  becomes  zero  at  a  gas  flow  of  297  Ib  per  second,  at  which  point  the 
draft  loss  due  to  friction  is  equal  to  the  theoretical  intensity.  The  draft- 
capacity  curve  corresponds  to  the  head-capacity  curve  of  centrifugal 
pump  characteristics  and  the  dynamic-head-capacity  curve  of  a  fan.  The 
point  of  maximum  draft  and  zero  capacity  is  called  shut-off  draft,  or  point 
of  impending  delivery,  and  corresponds  to  the  point  of  shut-off  head  of  a 
centrifugal  pump.  The  point  of  zero  draft  and  maximum  capacity  is 
called  the  wide  open  point  and  corresponds  to  the  wide  open  point  of  a 
centrifugal  pump.  A  set  of  operating  characteristics  may  be  developed 
for  any  size  chimney  operating  under  any  set  of  conditions  by  substituting 
the  proper  values  in  Equation  1  and  then  plotting  the  results  in  the 
manner  shown  in  Fig.  4. 

The  efficiency  of  a  natural  draft  chimney  is  the  thermodynamical  ratio 
of  the  energy  output  to  the  energy  input.  The  energy  output  is  the  total 

426 


CHAPTER  26 — CHIMNEYS  AND  DRAFT  CALCULATIONS 

work  done  by  the  chimney  in  moving  the  gases  and  corresponds  to  the 
water  horsepower  of  a  centrifugal  pump,  or  the  total  work  done  by  a  fan 
in  moving  the  gases.  The  energy  input  is  equal  to  the  theoretical  amount 
of  power  generated  by  the  chimney  and  corresponds  to  the  power  input 
of  the  driving  unit  of  a  centrifugal  pump  or  a  fan.  The  thermodynamical 
efficiency  is  given  by  the  equation: 


Et  = 


A\/H 


(2} 


where 


=  a  constant  depending  upon  the  temperature  of  the  gases,  the  atmospheric 
temperature,  the  elevation  of  the  plant,  and  the  density  and  specific  heat  of  the 
gases.  For  average  operating  conditions,  K&  —  0.0065. 


JU 

1.2 

1.1 
u, 

I0'9 

gQJB 

c. 
i^O.6 
£ 
QQ.5 
_g> 

TO  0.4 
1-03 
0.2 
0.1 
0 

*-  .— 

—  —  « 

—  — 

.0010 
.0009 
.0008 
.0007 
.0006 
.0005 
.0004 
.0003 
.0002 
.0001 
0 

"^ 

^ 

^ 

°*r» 

^ 

/ 

-^-* 

.  — 

X 

~^-^, 

-s^ 

,*< 

>^ 

^ 

\ 

\ 

X^ 

\ 

s 

\ 

/ 

\ 

v 

\ 

/ 

\ 

s. 

\ 

/ 

\ 

s 

^ 

/ 

\ 

\J 

/ 

\ 

\ 

^_ 

^\ 

)           30          60          90         120        150        180        210        240        270       300 
Amount  of  Gases  Flowing  and  Discharged,  Ib.  per  sec.,  W 

£ 

i.  !y 


FIG.  4.    TYPICAL  SET  OF  OPERATING  CHARACTERISTICS  OF  A  NATURAL  DRAFT  CHIMNEY 


Fig.  4  shows  the  variation  in  the  efficiency  of  the  chimney  under  con- 
sideration for  the  operating  conditions  noted.  This  curve  rises  from  zero 
at  shut-off  draft  to  a  maximum  for  a  certain  draft  and  its  corresponding 
capacity  and  then  drops  again  to  zero  at  the  wide  open  point.  The  point 
of  maximum  efficiency  is  located  by  the  point  on  the  draft-capacity  curve 
equal  to  two-thirds  of  the  theoretical  draft  intensity.  In  Example  1  the 
maximum  efficiency  is  at  an  available  draft  intensity  of  %  X  1.27  =  0.85 
in.  of  water  and  the  corresponding  capacity  of  175  Ib  per  second. 

The  efficiency  curve  of  a  natural  draft  chimney  corresponds  to  the 
efficiency  curves  of  a  centrifugal  pump  and  a  fan  and  serves  the  same 
general  use  in  that  it  locates  the  region  of  most  economical  operation.  In 
substituting  the  values  for  the  various  factors  in  Equation  1,  care  should 
be  exercised  that  the  selections  be  as  near  the  actual  conditions  as  is 
practically  possible.  The  following  notes  will  serve  as  a  guide  for  these 
selections : 

427 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


1.  The  barometric  pressure  varies  inversely  as  the  altitude  of  the  plant  above  sea  level. 
Fig.  5  gives  the  barometric  pressure  corresponding  to  various  elevations  as  computed 
from  the  equation: 


where 


62,737  log  - 


altitude  of  plant  above  sea  level,  feet. 


(3) 


In  general,  the  barometric  pressure  decreases  approximately  0.1  in.  of  mercury  per  100 
ft  increase  in  elevation. 

2.  The  unit  weight  of  a  cubic  foot  of  chimney  gases  at  0  F  and  sea  level  barometric 
pressure  is  given  by  the  equation  : 


Wc  =  0.131  CO*  +  0.095 


0.083 


(4) 


29 

^28 
*o 
c  27 

OJ 

=  26 
J-25 
'=  24 

|  23 

CO 

CD  22 
21 

?n 

\ 

^x 

N 

\ 

^»s_ 

^v 

. 

\ 

^x 

\ 

^ 

k^ 

-^ 

0         1000      2000      3000      4000      5000       6000      7000 
Corresponding  Altitude  above  Sea  Level,  ft 

FIG.  5.    RELATION  BETWEEN  BAROMETRIC  PRESSURE  AND  ALTITUDE 


In  this  equation  CO?,  0%  and  Ns  represent  the  percentages  of  the  parts  by  weight  of  the 
carbon  dioxide,  oxygen  and  nitrogen  content,  respectively,  of  the  gas  analysis.  For 
ordinary  operating  conditions,  the  value  of  W^  may  be  assumed  at  0.09. 

3.  The  atmospheric  temperature  is  the  actual  observed  temperature  of  the  outside  air 
at  the  time  the  analysis  of  the  operating  chimney  is  made.     The  mean  atmospheric 
temperature  in  the  temperate  zone  is  approximately  62  F. 

4.  The  chimney  gas  temperature  does  not  vary  appreciably  from  the  gas  temperature 
as  it  leaves  the  breeching  and  enters  the  chimney.    For  average  operating  conditions,  the 
chimney  gas  temperature  will  vary  between  500  F  and  650  F  except  in  the  case  when 
economizers  and  recuperators  are  used,  when  the  temperature  will  vary  between  300  F 
and  450  F.    If  a  chimney  has  been  properly  constructed,  properly  lined  and  has  no  air 
infiltration  due  to  open  joints,  the  temperature  of  the  gases  throughout  the  chimney  will 
not  differ  appreciably  from  the  foregoing  figures.    In  most  up-to-date  heating  plants,  the 
temperature  may  be  read  from  instruments  or  ascertained  from  a  pyrometer. 

5.  The  coefficient  of  friction  between  the  chimney  gases  and  a  sooted  surface  has  been 
found  to  be  approximately  0.016.    This  factor,  of  course,  will  be  much  less  for  a  new 
unlined  steel  stack  than  for  a  brick  or  brick-lined  chimney,  but  in  time  the  inside  surface 
of  all  chimneys  regardless  of  the  material  of  which  they  are  constructed  becomes  covered 
with  a  layer  of  soot  and  the  coefficient  of  friction  should  be  the  same  for  all  types  of 
chimneys. 

6.  The  length  of  the  friction  duct  is  the  vertical  distance  between  the  bottom  of  the 
breeching  opening  and  the  top  of  the  chimney.    Ordinarily  this  distance  is  approximately 
equal  to  the  height  of  the  chimney  above  the  grate  level. 

428 


CHAPTER  26 — CHIMNEYS  AND  DRAFT  CALCULATIONS 

7.  The  amount  of  gases  flowing  and  being  discharged  is,  of  course,  equal  to  the  amount 
of  gases  generated  in  the  combustion  chamber  of  the  boiler.  The  total  products  of 
combustion  may  be  computed  from  the  equation: 


W 


where 


CgGWtp 

3600 


Cg  =  pounds  of  fuel  burned  per  square  foot  of  grate  surface  per  hour. 
G  =  total  grate  surface  of  boilers,  square  feet, 
f/tp  =  total  weight  of  products  of  combustion  per  pound  of  fuel. 


(5) 


.001       .002        .003        .004       .005       .006       .007 
Available  Draft  per  Ft.  of  Height,  in.  of  Water 

FIG,  6.    CHIMNEY  PERFORMANCE  CHART 


Fig.  6  is  a  typical  chimney  performance  chart  giving  the  available  draft 
intensities  for  various  amounts  of  gases  flowing  and  sizes  of  ^  chimney. 
This  chart  is  based  on  an  atmospheric  temperature  of  62  F,  a  chimney  gas 
temperature  of  500  F,  a  unit  chimney  gas  weight  of  0.09  Ib  per  cubic  foot, 
sea  level  atmospheric  pressure,  a  coefficient  of  friction  of  0.016,  and  a 
friction  duct  length  equal  to  the  height  of  the  chimney  above  the  grate 
level.  These  curves  may  be  used  for  general  operating  conditions.  For 
specific  operating  conditions,  a  new  chart  should  be  constructed  from 
Equation  1. 

It  has  been  the  usual  custom,  and  still  is  to  a  lamentably  igreat  extent, 
to  select  the  required  size  of  a  natural  draft  chimney  from  a  table  of 

429 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

chimney  sizes  based  only  on  boiler  horsepowers.  After  the  ultimate 
horsepower  of  the  projected  plant  had  been  determined,  the  chimney  size 
in  the  table  corresponding  to  this  figure  was  then  selected  as  the  proper 
size  required.  Generally,  no  further  attempt  was  made  to  determine  if 
the  height  thus  selected  was  sufficient  to  help  create  the  required  draft 
demanded  by  the  entire  installation,  or  the  diameter  sufficiently  large  to 
enable  the  chimney  quickly,  efficiently,  and  economically  to  dispose  of  the 
gases.  Since  the  operating  characteristics  of  a  natural  draft  chimney  are 
similar  in  all  respects  to  those  of  a  centrifugal  pump,  or  a  centrifugal  fan, 
it  is  no  more  possible  to  select  a  proper  size  chimney  from  such  a  table, 
even  with  correction  factors  appended,  than  it  is  to  select  the  proper  size 
pump  from  tables  based  only  on  the  amount  of  water  to  be  delivered. 

DETERMINING  CHIMNEY  SIZES 

The  required  diameter  and  height  of  a  natural  draft  chimney  are  given 
by  the  following  equations: 

A- 


D  =  0.288  J    WTc  ('} 

1  BoWcV 

where 

H  —  required  height  of  chimney  above  grate  bar  level,  feet. 
D  =  required  minimum  diameter  of  chimney,  feet. 
V  =  chimney  gas  velocity,  feet  per  second. 

DT  =  total  required  draft  demanded  by  the  entire  installation  outside  of  the  chimney, 
inches  of  water. 

Equations  6  and  7  give  the  required  size  of  a  natural  draft  chimney  with 
all  of  the  operating  factors  taken  into  consideration.  Values  for  all  of  the 
factors  with  the  exception  of  the  chimney  gas  velocity  may  be  either 
observed  or  computed.  It  is,  of  course,  necessary  to  assume  an  arbitrary 
value  for  the  velocity  in  order  to  arrive  at  some  definite  size.  For  any  one 
set  of  operating  conditions  there  will  be  as  many  sizes  of  chimneys  as  there 
are  values  of  reasonable  velocities  to  assume.  Of  the  number  of  sizes 
corresponding  to  the  various  assumed  velocities,  there  is  one  size  which 
will  cost  least.  Since  the  cost  of  a  chimney  structure,  regardless  of  the 
kind  of  material  used  in  the  construction,  varies  as  the  volume  of  material 
in  the  structure,  the  cost  criterion  then  may  be  represented  by  the 
approximate  equation: 

Q  =  *tHD  (8) 

where 

Q  —  volume  of  material,  cubic  feet. 
/  =  average  wall  thickness,  feet. 

For  all  practical  purposes,  the  value  of  nt  may  be  taken  as  a  constant 
regardless  of  the  size  of  the  structure.  Hence,  in  general,  the  voluine,  and 
consequently  the  cost,  of  a  chimney  structure  may  be  based  on  the  factor 

430 


CHAPTER  26 — CHIMNEYS  AND  DRAFT  CALCULATIONS 


HD  as  a  criterion.  Therefore,  the  value  of  the  chimney  gas  velocity  which 
will  result  in  the  least  value  of  HD  for  any  one  set  of  operating  con- 
ditions will  produce  a  structure  which  will  be  the  most  economical  to  use, 
because  its  cost  will  be  least. 

The  problem  at  hand  is  to  deduce  an  equation  for  the  chimney  gas 
velocity  which  will  result  in  a  combination  of  a  height  and  a  diameter 
whose  product  HD  will  be  least.  The  solution  is  obtained  by  equating  the 


200 
190 
180 
170 
^.160 
§150 

&140 
£ 
•g-130 

|120 

|no 

"glOO 
|  90 
1    80 

1    7° 
I    60 
|    50 
|    40 
30 
20 
10 

°( 

Height  of  Chimney,  ft. 
D    25    50    7510012515017520022525027530032535037 

2.0 
1.9 
1.8 
1.7 
1.6 
1.5 

W    , 
13?- 
12  | 
l,o 

1.0  d- 

0.9  Q 
0.8  I 

ay  I 

0.6 
0.5 
0.4 
0.3 
02 
0.1 

5° 

/ 

/ 

/ 

/ 

/ 

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' 

/ 

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v^ 

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£f 

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7 

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/ 

/ 

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P 

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,_--—  • 

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,> 

)      1      2     3     4      5     6      7     8     9    10    11    12    13    14   1 
Diameter  of  Chimney,  ft. 

FIG.  7.    ECONOMICAL  CHIMNEY  SIZES 

product  of  Equations  3  and  4  to  HD,  differentiating  this  product  with 
respect  to  V  and  equating  the  resulting  expression  to  zero.  This  pro- 
cedure results  in  the  following  expression : 


2/5 


(9) 


where  Fe  ~  economical  chimney  gas  velocity,  feet  per  second. 

Equation  9  gives  the  economical  velocity  of  the  chimney  gases  for  any 
set  of 'Operating  conditions,  and  represents  the  velocity  which  will  result  in 
a  chimney  the  size  of  which  will  cost  less  than  that  of  any  other  size  as 
determined  by  any  other  velocity  for  the  same  operating  conditions. 
After  the  value  of  the  economical  velocity  has  been  determined,  the 

431 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

corresponding  height  and  diameter  can  then  be  determined  from  Equa- 
tions 6  and  7,  respectively,  and  the  economical  size  will  then  be  attained. 
Equations  6,  7  and  9  may  be  simplified  considerably  for  average  operating 
conditions  in  an  average  size  steam  plant  by  assuming  the  following 
conditions: 

Average  chimney  gas  temperature,  500  F Tc  =  960 

Mean  atmospheric  temperature,  62  F T0  =  522 

Average  coefficient  of  friction,  0.016 ./  =  0.016 

Average  chimney  gas  density,  0.09 Wc  =  0.09 

Sea  level  elevation  with  barometer  of  29.92 B0  -  29.92 

Substituting  these  values  in  Equations  9,  7  and  6,  respectively,  and 
reducing: 

7e 
D 
H  =  190Dr  (12) 

Fig.  7  gives  the  economical  chimney  sizes  for  various  amounts  of  gases 
flowing  and  for  required  draft  intensities  as  computed  from  Equations  10, 
11  and  12.  They  are  based  on  the  operating  factors  used  in  reducing 
Equations  6,  7  and  9  to  their  simpler  form.  The  sizes  shown  by  the 
curves  in  the  chart  should  be  used  for  general  operating  conditions  only, 
or  for  installations  where  the  required  data  necessary  for  an  exact  deter- 
mination are  difficult  or  impossible  to  secure.  Whenever  it  is  possible  to 
secure  accurate  data,  or  the  anticipated  operating  conditions  are  fairly 
well  known,  the  required  size  should  be  determined  from  Equations  6, 
7  and  9.  The  recommended  minimum  inside  dimensions  and  heights  of 
chimneys  for  small  and  medium  size  installations  are  given  in  Table  1. 

GENERAL  EQUATION 

The  general  draft  equation  for  a  steam  producing  plant  may  be  stated 
as  follows: 

Dt  -  hf  =  to  -f  AB  +  h*d  4-  he  4-  hBr  +  hv  +  ho  -1-  k&  +  £R  (13) 

where 

Dt  =  theoretical  draft  intensity  created  by  pressure  transformer,  inches  of  water. 
h{  —  draft  loss  due  to  friction  in  pressure  transformer,  inches  of  water, 
/rp  =  draft  loss  through  the  fuel  bed,  inches  of  water. 
/ZB  ==  draft  loss  through  the  boiler  and  setting,  inches  of  water. 
h%T  =  draft  loss  through  the  breeching,  inches  of  water. 
hv  =  draft  loss  due  to  velocity,  inches  of  water. 
^Bd  =  draft  loss  due  to  bends,  inches  of  water. 
he  =  draft  loss  due  to  contraction  of  opening,  inches  of  water. 
ho  =  draft  loss  due  to  enlargement  of  opening,  inches  of  water. 
&E  —  draft  loss  through  the  economizer,  inches  of  water. 
&R  =  draft  loss  through  recuperators,  regenerators,  or  air  heaters,  inches  of  water. 

The  left  hand  member  of  Equation  13  represents  the  total  amount  of 
available  draft  created  by  the  pressure  transformer,  that  is,  the  natural 

432 


CHAPTER  26 — CHIMNEYS  AND  DRAFT  CALCULATIONS 


TABLE  1. 


RECOMMENDED  MINIMUM  CHIMNEY  SIZES  FOR 
HEATING  BOILERS  AND  FURNACES  a 


H.-_ 

RECTANGULAR  FLUE 

ROUND  FLUE 

WARM  Am 
FURNACE 
CAPACITY 
IN  SQ  Lv. 
OF  LEADER 

STEAM 
BOILER 
CAPACITY 
SQ  FT 
OP  RADI- 
ATION 

OT 

WATER 
HEATER 
CAPACITY 
SQFT 
OF  RADI- 
ATION 

NOMINAL 
DIMEN- 
SIONS OP 
FIRE  CLAY 
LINING 
IN  INCHES 

HEIGHT 
IN   FT 
ABOVE 
GRATE 

Actual 
Inside 
Dimensions 
of  Fire  Clay 

t  Lining 

Actual 
Area 
Sq  In. 

Inside 
Diam- 
eter of 
Lining 
in 

Actual 
Area 
Sq  In. 

in  Inches 

Inches 

790 

590 

973 

8j^x  13 

7      X  ll/^ 

81 

35 

1000 

690 

1,140 

10 

79 

900 

1,490 

13x13 

113^  x  llj^ 

127 

900 

1,490 

83^x18 

65^  x  16/<£ 

110 

1,100 

1,820 

12 

113 

40 

1,700 

2,800 

13x18 

llj^x  16J4 

183 

1,940 

3,200 

15 

177 

2,130 

3,520 

18x18 

15^x15% 

248 

2,480 

4,090 

20x20 

298 

45 

3,150 

5,200 

18 

254 

50 

4,300 

7,100 

20 

314 

4,600 

7,590 

20x24 

17x21 

357 

5,000 

8,250 

24x24 

21x21 

441 

55 

5,570 

9,190 

24x24t> 

576 

60 

5,580 

9,200 

22 

380 

6,980 

11,500 

24 

452 

65 

7,270 

12,000 

24x28t> 

672 

8,700 

14,400 

28x28^ 

784 

9,380 

15,500 

27 

573 

10,150 

16,750 

30x30^ 

900 

10,470 

17,250 

28x32b 

896 

aThis  table  is  taken  from  the  A.S.H.V.E.  Code  of  Minimum  Requirements  for  the  Heating  and  Venti- 
lation of  Buildings  (Edition  of  1929). 

bDimensions  are  for  unlined  rectangular  flues. 


draft  chimney,  Venturi  chimney,  or  fan,  and  is  equal  to  the  theoretical 
intensity  less  the  internal  losses  incidental  to  operation.  The  right  hand 
member  represents  the  sum  of  all  of  the  various  losses  of  draft  throughout 
the  entire  boiler  plant  installation  outside  of  the  pressure  transformer 
itself.  The  left  hand  member  expresses  the  available  intensity  and  is 
analogous  to  the  head  developed  by  a  centrifugal  pump  in  a  water  works 
system,  while  the  right  hand  member  expresses  the  required  draft  in- 
tensity and  is  analogous  to  the  total  dynamic  head  in  a  water  works 
system.  For  a  general  circulation  of  gases 

£>a  -  Dr  (14) 


where 


DT 


available  draft  intensity,  inches  of  water. 
>  required  draft,  inches  of  water. 


The  draft  loss  through  the  fuel  bed  (Ap)>  or  the  amount  of  draft  required  to 
effect  a  given  or  required  rate  of  combustion,  varies  between  wide  limits 
and  represents  the  greater  portion  of  the  required  draft.  In  coal-fired 
installations,  the  draft  loss  through  the  fuel  bed  is  dependent  upon  the 
following  factors:  (1)  character  and  condition  of  the  fuel,  clean  or  dirty; 
(2)  percentage  of  ash  in  the  fuel;  (3)  volume  of  interstices  in  the  fuel  bed, , 

433 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

coarseness  of  fuel;  (4)  thickness  of  the  fuel  bed,  rate  of  combustion; 
(5)  type  of  grate  or  stoker  used;  (6)  efficiency  of  combustion. 

There  is  a  certain  intensity  of  draft  with  which  the  best  results  will  be 
obtained  for  every  kind  of  coal  arid  rate  of  combustion.  Fig.  8  gives  the 
intensity  of  draft,  or  the  vacuum  in  the  combustion  chamber  required  to 
burn  various  kinds  of  coal  at  various  rates  of  combustion.  Expressed  in 
other  words,  these  curves  represent  the  amount  of  draft  required  to  force 
the  necessary  amount  of  air  through  the  fuel  bed  in  order  to  effect  various 
rates  of  combustion.  It  will  be  noted  that  the  amount  of  draft  increases 
as  the  percentage  of  volatile  matter  diminishes,  being  comparatively  low 
for  the  lower  grades  of  bituminous  coals  and  highest  for  the  high  grades 
and  small  sizes  of  anthracites.  Also,  when  the  interstices  of  the  coal  are 
large  and  the  particles  are  not  well  broken  up,  as  with  bituminous  coals, 


05  10          15          £0          25         30         35         40         45 

Pounds  of  Coal  Burned  per  5<7|.Ft  of  Grctte  Surface  per  Hr: 

FIG.  8.    DRAFT  REQUIRED  AT  DIFFERENT  RATES  OF  COMBUSTION 
FOR  VARIOUS  KINDS  OF  COAL 


much  less  draft  is  required  than  when  the  particles  are  small  and  are  well 
broken  up,  as  with  bituminous  slack  and  the  small  sizes  of  anthracites.  In 
general,  the  draft  loss  through  the  fuel  bed  increases  as:  (1)  the  per- 
centage of  volatile  matter  diminishes;  (2)  the  percentage  of  fixed  carbon 
increases;  (3)  the  thickness  of  the  bed  increases;  (4)  the  percentage  of  ash 
increases;  (5)  the  volume  of  the  interstices  diminishes. 

In  making  the  preliminary  assumptions  for  the  draft  loss  through  the 
fuel  bed,  due  allowances  should  be  made  for  a  possible  future  change  in 
the  grade  of  fuel  to  be  burned  and  also  in  the  rate  of  combustion.  A  value 
should  be  selected  for  this  loss  which  will  represent  not  only  the  highest 
rate  of  combustion  which  will  be  encountered,  but  also  the  grade  of  coal 
which  has  the  greatest  resistance  through  the  fuel  bed  and  which  may  be 
burned  at  a  later  date. 

In  powdered-fuel  and  oil-fired  installations,  there  will  be  no  draft  loss 

434 


CHAPTER  26  —  CHIMNEYS  AND  DRAFT  CALCULATIONS 

through  the  fuel  bed  since  there  is  none  and,  consequently,  this  factor 
becomes  zero  in  the  general  draft  equation.  All  other  factors  being 
constant,  the  height  of  the  chimney  in  installations  of  this  character  will 
be  less  than  the  height  in  coal-fired  Installations,  and  in  the  case  of  me- 
chanical draft  installations  the  driving  units  need  not  be  as  large  since  the 
head  against  which  the  fan  is  to  operate  is  not  as  great  in  the  former  as 
in  the  latter. 

The  draft  loss  through  the  boiler  and  setting  (h%)  also  varies  between  wide 
limits  and,  in  general,  depends  upon  the  following  factors: 

1.  Type  of  boiler.  5.  Arrangement  of  baffles. 

2.  Size  of  boiler.  6.  Type  of  grate. 

3.  Rate  of  operation.  7.  Design  of  brickwork  setting. 

4.  Arrangement  of  tubes.  8.  Excess  air  admitted. 

9.  Location  of  entrance  into  breeching. 

Curves  showing  the  draft  loss  through  the  boiler  are  usually  based  on 
the  load  or  quantity  of  gases  passing  through  the  boiler,  expressed  in 
terms  of  percentage  of  normal  rate  of  operation.  Owing  to  the  great 
variety  of  boilers  of  different  designs  and  the  various  schemes  of  baffling, 
it  is  impossible  to  group  together  a  set  of  curves  for  the  draft  loss  through 
the  boiler  which  may  even  be  used  generally.  It  is  therefore  necessary  to 
secure  this  information  from  the  manufacturer  of  the  particular  type  of 
boiler  and  baffle  arrangement  under  consideration. 

When  a  boiler  is  installed  and  in  operation,  the  draft  loss  depends  upon 
the  amount  of  gases  flowing  through  it.  This,  in  turn,  depends  upon  the 
proportion  of  excess  air  admitted  for  combustion.  The  amount  of  excess 
air  is  measured  by  the  C02  content;  the  less  the  amount  of  C02,  the 
greater  the  amount  of  excess  air  and  hence  the  greater  the  draft  loss. 

The  loss  of  draft  through  the  boiler  will  vary  directly  as  the  size  of  the 
boiler  and  the  length  of  the  gas  passages  within.  The  loss  also  varies  as 
the  number  of  tubes  high,  but  not  in  a  direct  ratio  inasmuch  as  the  loss 
due  to  the  reversal  of  flow  at  the  ends  of  the  baffles  remains  constant 
regardless  of  the  height  of  the  boiler.  The  arrangement  of  the  tubes, 
whether  the  gases  flow  parallel  to  or  at  right  angles  to  the  tubes,  has  an 
appreciable  effect  on  the  loss.  The  arrangement  of  the  baffles  influences 
the  draft  loss  greatly,  the  loss  through  a  boiler  with  five  passes  being 
greater  than  the  loss  through  one  of  three  or  four  passes.  A  poor  design 
arid  a  rough  condition  of  the  brickwork  will  increase  the  loss  greatly, 
whereas  a  proper  design  and  a  smooth  condition  will  keep  the  loss  at  a 
minimum.  The  loss  through  the  boiler  will  be  less  when  the  breeching 
entrance  is  located  at  or  near  the  top  of  the  boiler  than  when  it  is  located 
at  or  near  the  bottom  since  the  gases  have  a  shorter  distance  to  travel 
in  the  former  instance. 

1  The  draft  loss  through  the  breeching  (/br)  is  given  by  the  general 
equation  : 


435 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

where 

W  =  the  amount  of  gases  flowing,  pounds  per  second. 
Tc  =  absolute  temperature  of  breeching  gases,  degrees  Fahrenheit. 
/  =  coefficient  of  friction. 
L  —  length  of  breeching,  feet. 
A  —  area  of  breeching,  square  feet. 

BQ  =  atmospheric  pressure  corresponding  to  altitude,  inches  of  mercury. 
Wc  =  weight  of  a  cubic  foot  of  breeching  gases  at  0  F  and  sea  level  atmospheric 

pressure,  pounds  per  cubic  foot. 
Cbr  =  hydraulic  radius  of  breeching  section. 

It  has  been  the  general  custom  to  lump  off  the  intensity  of  the  breeching 
loss  at  0.10  in.  of  water  per  100  ft  of  breeching  length  regardless  of  its  size 
or  shape  or  the  amount  and  temperature  of  the  gases  flowing  through  it. 
This  practice  is  hazardous  and  has  no  more  foundation  in  fact  than  that  of 
determining  the  friction  head  in  a  water  works  system  without  taking 
into  consideration  the  size  of  the  pipe  or  the  amount  of  water  flowing 
through  it.  When  the  length  of  the  breeching  is  relatively  short,  any 
variation  in  any  one  of  the  factors  in  the  equation  will  have  no  appreciable 
effect  on  the  draft  loss.  However,  when  the  breeching  is  relatively  long, 
the  draft  loss  is  affected  greatly  by  the  various  factors,  particularly  by  the 
size  and  shape  as  well  as  by  the  weight  of  gases  flowing. 

The  draft  loss  due  to  velocity  (hv)  is  given  by  the  equation 


and  represents  the  amount  of  draft  required  to  accelerate  the  gases  from 
zero  velocity  to  the  velocity  at  which  the  gases  are  flowing,  or  in  other 
words,  from  a  static  gas  condition  of  zero  flow  to  the  amount  of  gases 
flowing  throughout  the  installation.  This  loss  corresponds  to  the  velocity 
head  in  water  works  systems. 

The  draft  loss  due  to  bends  (&sd)  is  equivalent  to  the  loss  due  to  the 
velocity  head  for  a  90-deg  bend.  In  changing  direction  of  flow,  the  gas 
velocity  decreases  to  zero  with  a  loss  of  velocity  head  and  then  increases 
to  its  proper  value  at  the  expense  of  a  loss  in  pressure  head,  the  net  result 
being  a  loss  in  pressure  head  equal  to  the  velocity  head  at  the  bend. 
This  loss  is  given  by  the  equation  : 

,  0.000194  W*TC  /.~v 


The  friction  at  a  right-angle  bend  is  sometimes  expressed  as  the 
equivalent  of  a  straight  length  of  flue  of  a  certain  length  for  a  certain 
diameter,  similar  to  the  procedure  used  in  estimating  the  loss  due  to 
bends  in  piping  systems  conducting  water.  Most  flues,  however,  par- 
ticularly breechings,  are  built  square  or  rectangular  in  section  and  no 
general  equation  based  on  the  shape  of  the  flue  can  be  conveniently 
expressed. 

The  draft  loss  due  to  sudden  contraction  of  an  area  (he)  is  given  by  the 
equation  : 

436 


CHAPTER  26  —  CHIMNEYS  AND  DRAFT  CALCULATIONS 


. 

he  =  ,,2R  T,7  (18) 


Kc  =  coefficient  of  sudden  contraction  based  on  -—  ,  the  ratio  of  the  areas  of  the 

•"•i 
smaller  to  the  larger  section. 

As  =  area  of  the  smaller  section. 

When  the  flue  or  passage  through  which  the  gases  flow  is  suddenly 
contracted,  a  considerable  portion  of  the  static  head  in  the  larger  section 
is  converted  into  velocity  head  and  a  draft  loss  of  some  consequence,  par- 
ticularly in  a  short  breeching,  takes  place.  A  sudden  contraction  should 
always  be  avoided  where  possible.  At  times,  however,  due  to  obstruc- 
tions or  limited  head-room,  it  is  necessary  to  alter  the  size  of  the  breeching, 
but  a  sudden  contraction  may  be  avoided  by  gradually  decreasing  the 
area  over  a  length  of  several  feet. 

The  draft  loss  due  to  a  sudden  enlargement  of  an  area  (ho)  is  given  by  the 
equation  : 

0.000194£0  J7*rc 

h°  -          AlBoW*  -  (19) 

where 

A 

KO  =  coefficient  of  sudden  enlargement  based  on  -~,  the  ratio  of  the  areas  of  the 

A\ 

smaller  to  the  larger  section. 

When  the  flue  or  passage  through  which  the  gases  flow  is  suddenly 
enlarged,  a  portion  of  the  velocity  head  is  converted  into  static  head  in  the 
larger  section  and,  like  the  loss  due  to  sudden  contraction,  a  loss  of  some 
consequence,  particularly  in  short  breechings,  takes  place.  A  sudden 
enlargement  in  a  breeching  may  be  avoided  by  gradually  increasing  the 
area  over  a  length  of  several  feet.  In  large  masonry  chimneys,  the  area  of 
the  flue  at  the  region  of  the  breeching  entrance  is  considerably  larger 
than  the  area  of  the  breeching  at  the  chimney,  and  a  sudden  enlargement 
exists. 

The  draft  loss  through  the  economizer  (fe)  should  be  obtained  from  the 
manufacturer  but  for  general  purposes  it  may  be  computed  from  the 
following  general  equation: 


te  -      "c  (20) 

10™ 

where 

Wn  =  pounds  of  gases  flowing  per  hour  per  linear  foot  of  pipe  in  each  economizer 

section. 
N  =s  number  of  economizer  sections. 

An  economizer  in  a  steam  plant  affects  the  draft  in  two  ways,  (1)  it 
offers  a  resistance  to  the  flow  of  gases,  and  (2)  it  lowers  the  average 
chimney  gas  temperature,  thereby  decreasing  the  available  intensity.  In 
the  case  of  a  natural  draft  installation,  both  of  these  factors  result  in  a 
relative  increase  in  the  height  of  the  chimney  and,  in  the  case  of  a  large 

437 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

plant,  they  may  add  as  much  as  20  or  30  ft  to  the  height.  The  decrease 
in  the  temperature  of  the  gases  after  they  have  passed  through  the 
economizer  has  an  extremely  important  effect  on  the  performance  of  a 
natural  draft  chimney  and  also  upon  the  performance  of  a  fan. 

CONSTRUCTION  DETAILS 

For  general  data  on  the  construction  of  chimneys  reference  should  be 
made  to  the  Standard  Ordinance  for  Chimney  Construction  of  the 
National  Board  of  Fire  Underwriters.  Briefly  summarized,  these  provisions 
are  as  follows  for  heating  boilers  and  furnaces : 

The  construction,  location,  height  and  area  of  the  chimney  to  which  a  heating  boiler 
or  warm-air  furnace  is  connected  affect  the  operation  of  the  entire  heating  system.  Most 
residence  chimneys  are  built  of  brick  and  may  be  either  lined  or  unlined,  but  in  either 
case  the  walls  must  be  air-tight  and  there  should  be  only  one  smoke  opening  into  the 
chimney.  Cleanout,  if  provided,  must  be  absolutely  air-tight  when  closed. 

The  walls  of  brick  chimneys  shall  be  not  less  than  3J£  in.  thick  (width  of  a  standard 
size  brick)  and  shall  be  lined  with  fire-clay  flue  lining.  Fire-clay  flue  linings  shall  be 
manufactured  from  suitable  refractory  clay,  either  natural  or  compounded,  and  shall 
be  adapted  to  withstand  high  temperatures  and  the  action  of  flue  gases.  They  shall  be 
of  standard  commercial  thickness,  but  not  less  than  %  in-  All  fire-clay  flue  linings  shall 
meet  the  standard  specification  of  the  Eastern  Clay  Products  Association.  The  flue 
sections  shall  be  set  in  special  mortar,  and  shall  have  the  joints  struck  smooth  on  £he 
inside.  The  masonry  shall  be  built  around  each  section  of  lining  as  it  is  placed,  and  all 
spaces  between  masonry  and  linings  shall  be  completely  filled  with  mortar.  No  broken 
flue  lining  shall  be  used.  Flue  lining  shall  start  at  least  4  in.  below  the  bottom  of  smoke- 
pipe  intakes  of  flues,  and  shall  be  continued  the  entire  heights  of  the  flues  and  project 
at  least  4  in.  above  the  chimney  top  to  allow  for  a  2  in.  projection  of  lining.  The  wash  or 
splay  shall  be  formed  of  a  rich- cement  mortar.  To  improve  the  draft  the  wash  surface 
should  be  concave  wherever  practical. 

Flue  lining  may  be  omitted  in  brick  chimneys,  provided  the  walls  of  the  chimneys 
are  not  less  than  8  in.  thick,  and  that  the  inner  course  shall  be  a  refractory  clay  brick. 
All  brickwork  shall  be  laid  in  spread  mortar,  with  all  joints  push-filled.  Exposed  joints 
both  inside  and  outside  shall  be  struck  smooth.  No  plaster  lining  shall  be  permitted. 

Chimneys  shall  extend  at  least  3  ft  above  flat  roofs  and  2  ft  above  the  ridges  of  peak 
roofs  when  such  flat  roofs  or  peaks  are  within  30  ft  of  the  chimney.  The  chimney 
shall  be  high  enough  so  that  the  wind  from  any  direction  shall  not  strike  the  top  of  the 
chimney  from  an  angle  above  the  horizontal.  The  chimney  shall  be  properly  capped  with 
stone,  terra  cotta,  concrete,  cast-iron,  or  other  approved  material;  but  no  such  cap 
or  coping  shall  decrease  the  flue  area. 

There  shall  be  but  one  connection  to  the  flue  to  which  the  boiler  or  furnace  smoke- 
pipe  is  attached.  The  boiler  or  furnace  smoke-pipe  shall  be  thoroughly  grouted  into  the 
chimney  and  shall  not  project  beyond  the  inner  surface  of  the  flue  lining. 

The  size  or  area  of  flue  lining  or  of  brick  flue  for  warm-air  furnaces  depends  on  height 
of  chimney  and  capacity  of  heating  system.  For  chimneys  not  less  than  35  ft  in  height 
above  grate  line,  the  net  internal  dimensions  of  lining  should  be  at  least  7x  11 J^  in. 
for  a  total  leader  pipe  area  up  to  790  sq  in.  Above  790  and  up  to  1,000  sq  in.  of  leader 
pipe  area  the  lining  should  be  at  least  11  Ji  x  11 M  in.  inside.  In  case  of  brick  flues  not 
less  than  35  ft  in  height  with  no  linings,  the  internal  dimensions  should  be  at  Isast 
8  x  12  in.  up  to  790  sq  in.  of  leader  area,  and  at  least  12  x  12  in.  for  leader  capacities  up  to 
1,000  sq  in.  Chimneys  under  35  ft  in  height  are  unsatisfactory  in  operation  and  hence 
should  be  avoided.  .  N ' 

CHIMNEYS  FOR  CAS  HEATING 

The  burning  of  gas  differs  from  the  burning  of  coal  in  that  the  force 
which  supplies  the  air  for  combustion  of  the  gas  comes  largely  from  the 
pressure  of  the  gas  in  the  supply  pipe,  whereas  air  is  supplied  to  a  bedtoi 

438 


CHAPTER  26 — CHIMNEYS  AND  DRAFT  CALCULATIONS 

burning  coal  by  the  force  of  the  chimney  draft.  If,  with  a  coal-burning 
boiler,  the  draft  is  poor,  or  if  the  chimney  is  stopped,  the  fire  is  smothered 
and  the  combustion  rate  reduced.  In  a  gas  boiler  or  furnace  such  a 
condition  would  interfere  with  the  combustion  of  the  gas,  but  the  gas 
would  continue  to  pass  to  the  burners  and  the  resulting  incomplete  com- 
bustion would  produce  a  dangerous  condition.  In  order  to  prevent  incom- 
plete combustion  from  insufficient  draft,  all  gas-fired  boilers  and  furnaces 
should  have  a  back-draft  diverter  in  the  flue  connection  to  the  chimney. 

A  study  of  a  typical  back-draft  diverter  (Fig.  9)  shows  that  partial'  or 
complete  chimney  stoppage  will  merely  cause  some  of  the  products  of 
combustion  to  be  vented  out  into  the  boiler  room,  but  will  not  interfere 
with  combustion.  In  fact,  gas-designed  appliances  must  perform  safely 


--r 


-h, 


FIG.  9.    TYPICAL  BACK-DRAFT  DIVERTER 


under  such  a  condition  to  be  approved  by  the  American  Gas  Association 
Laboratory.  Other  functions  of  the  back-draft  diverter  are  to  protect  the 
burner  and  pilot  from  the  effects  of  down-drafts,  and  to  neutralize  the 
effects  of  variable  chimney  drafts,  thus  maintaining  the  appliance  ef- 
ficiency at  a  substantially  constant  value.  Converted  boilers  or  furnaces, 
as  well  as  gas-designed  appliances,  should  be  provided  with  back-draft 
diverter  s. 

As  is  the  case  with  the  complete  combustion  of  almost  all  fuels,  the 
products  of  combustion  for  gas  are  carbon  dioxide  (CO 2)  and  water  vapor 
with  just  a  trace  of  sulphur  trioxide  (503).  Sulphur  usually  burns  to  the 
trioxide  in  the  presence  of  an  iron  oxide  catalyst.  The  volume  of  water 
vapor  in  the  flue  products  is  about  twice  the  volume  of  the  carbon  dioxide 
when  coke  oven  or  natural  gas  is  burned.  Because  of  the  large  quantity 
of  water  vapor  which  is  formed  by  the  burning  of  gas,  it  is  quite  important 
that  all  gas-fired  central  heating  plants  be  connected  to  a  chimney  having 
a  good  draft.  Lack  of  chimney  draft  causes  stagnation  of  the  proiducts  of 
combustion  in  the  chimney  and  results  in  the  condensation  of  a  large 
amount  of  the  water  vapor.  A  good  chimney  draft  draws  air  into  the 
chimney  through  the  openings  in  the  back-draft  diverter,  lowers  the  dew 
point  of  the  mixture,  and  reduces  the  tendency  of  the  water  vapor  to 
condense.  ; 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

A  chimney  for  a  gas-fired  boiler  or  furnace  should  be  constructed  in 
accordance  with  the  principles  applicable  to  other  boilers.  Where  the 
wall  forming  a  smoke  flue  is  made  up  of  less  than  an  8-in.  thickness  of 
brick,  concrete,  or  stone,  a  burnt  fire  clay  flue  tile  lining  should  be  used. 
Care  should  be  used  that  the  lengths  of  flue  tile  meet  properly  with  no 
openings  at  the  joints.  Cement  mortar  should  be  used  for  the  entire 
chimney. 

TABLE  2.    MINIMUM  ROUND  CHIMNEY  DIAMETERS  FOR  GAS  APPLIANCES  (INCHES) 


HEIGHT  or 


GAS  CONSUMPTION  JN  THOUSANDS  OP  BTTT  PER  HOUR 


UHIMNBY 

FEET 

100 

200 

300 

400 

500 

750 

1000 

1500 

2000 

20 

4.50 

5.70 

6.60 

7.30 

8.00 

9.40 

10.50 

12.35 

13.85 

40 

4.25 

5.50 

6.40 

7.10 

7.80 

9.15 

10.25 

12.10 

13.55 

60 

4.10 

5.35 

6.20 

6.90 

7.60 

8.90 

10.00 

11.85 

13.25 

80 

4.00 

5.20 

6.00 

6.70 

7.35 

8.65 

9.75 

11.50 

12.85 

100 

3.90 

5.00 

5.90 

6.50 

7.20 

8.40 

9.40 

11.00 

12.40 

Table  2  gives  the  minimum  cross-sectional  diameters  of  round  chim- 
neys (in  inches)  for  various  amounts  of  heat  supplied  to  the  appliance, 
and  for  various  chimney  heights.  This  is  in  accordance  with  American 
Gas  Association  recommendations. 

The  flue  connections  from  a  gas-fired  boiler  or  furnace  to  the  chimney 
should  be  of  a  non-corrosive  material.  In  localities  where  the  price  of 
gas  requires  the  use  of  highly  efficient  appliances,  the  material  used  for 
the  flue  connection  not  only  should  be  resistant  to  the  corrosion  of  water, 
but  should  resist  the  corrosion  of  dilute  solutions  of  sulphur  trioxide  in 
water.  Sheet  aluminum,  as  well  as  some  other  materials,  seems  to  serve 
this  purpose  very  well. 


PROBLEMS  IN  PRACTICE 

1  •  What  is  draft? 

Draft  is  an  unbalanced  pressure  between  the  atmosphere  and  the  passages  in  the  ap- 
paratus or  construction  through  which  the  gases  flow. 

2  •  What  two  kinds  of  draft  need  be  considered? 

Natural  draft  caused  by  temperature  differences,  and  artificial  draft  caused  by  me- 
chanical forcing. 

3  •  What  is  the  effective  height  of  a  chimney? 

The  height  from  the  grate  level  to  the  top  of  the  chimney  is  the  effective  height  in  pro- 
ducing natural  draft. 

4  •  What  dual  purpose  does  a  tall  chimney  fulfill? 

A  tall  chimney  primarily  creates  the  necessary  draft  to  move  the  air  required  for  the 
combustion  process  and  to  move  the  products  of  combustion,  and  secondarily  it  dis- 
charges the  gases  at  a  high  elevation  to  prevent  them  from  becoming  a  nuisance. 

440 


CHAPTER  26 — CHIMNEYS  AND  DRAFT  CALCULATIONS 

5  •  a.  Name  the  principal  advantages  of  natural  draft. 

b.  Name  the  principal  disadvantages  of  natural  draft. 

a.  Simplicity,  reliability,  freedom  from  mechanical  parts,  low  cost  of  maintenance, 
relatively  long  life,  relatively  low  depreciation,  operation  with  no  power  requirement. 

i>.  Lack  of  flexibility,  irregularity,  dependence  on  surroundings,  susceptibility  to  tem- 
perature changes. 

6  •  How  is  mechanical  draft  created? 

By  forced  draft,  by  induced-draft  fans,  or  by  a  Venturi  chimney. 

7  •  Distinguish  between  theoretical  and  available  draft. 

Theoretical  draft  is  the  difference  in  pressure  inside  and  outside  the  base  of  a  chimney 
when  it  is  under  operating  temperatures  but  when  there  are  no  gases  flowing.  Available 
draft  is  less  than  theoretical  draft  by  the  friction  loss  due  to  the  flow  of  gases  through 
the  chimney. 

8  •  Explain  the  term  efficiency  of  a  natural  draft  chimney. 

The  efficiency  of  a  chimney  is  the  ratio  of  the  work  it  does  in  moving  gases  to  the  theo- 
retical amount  of  power  it  generates. 

9  •  How  is  the  available  draft  used  in  a  heating  plant? 

The  available  draft  at  the  base  of  the  chimney  is  used  to  overcome  the  loss  in  pressure 
through  the  grate,  the  fuel  bed,  the  boiler  passes,  the  breeching,  and  the  chimney. 

10  •  What  are  some  of  the  factors  that  influence  the  draft  loss  through  the  fuel 
bed? 

Uniformity  and  size  of  coal,  the  amount  of  ash  mixed  with  the  fuel  on  the  grate,  thickness 
of  fuel  bed,  rate  of  combustion,  amount  of  air  supply  as  related  to  the  coal  burning  rate. 

11  •  How  does  the  volatile  matter  content  affect  the  draft  loss  through  the  fuel 
bed? 

The  higher  the  volatile  content  and  the  lower  the  fixed  carbon  content,  the  lower  the 
draft  loss. 

12  •  In  what  cases  will  there  be  no  fuel  bed  draft  loss? 

In  oil,  gas,  and  powdered  fuel  firing  the  fuel  is  mixed  and  burned  in  suspension;  con- 
sequently, no  measurable  resistance  is  encountered  in  the  combustion  zone. 

13  •  Is  it  possible  to  state  an  average  value  for  the  draft  loss  through  a  boiler 
and  its  setting? 

No.  The  draft  loss  varies  widely  and  depends  on  many  factors  such  as  the  size  and  type 
of  gas  passageways.  The  manufacturer  is  usually  able  to  supply  such  information. 

14  •  Of  what  significance  is  the  CO2  content  of  stack  gases  in  establishing 
draft  loss? 

The  CO2  content  of  the  exit  gases  is  a  measure  of  the  completeness  of  the  combustion  and 
the  amount  of  excess  air  supplied.  Low  CO*  indicates  a  high  excess  of  air  and  hence 
a  high  draft  loss. 

15  •  "What  two  effects  does  an  economizer  have  on  the  draft  loss? 

An  economizer  offers  resistance  to  the  flow  of  gases  over  the  added  surfaces;  it  lowers  the 
temperature  of  the  gases  going  to  the  chimney  and  therefore  decreases  the  available 
draft.  This  decrease  often  necessitates  the  addition  of  forced  draft. 

16  •  What  main  provisions  should  be  considered  in  good  chimney  construction? 

Chimneys  should  be  air-tight  and  connected  to  only  one  smoke  opening.  The  chimney 
top  should  be  high  enough  above  surroundings  so  the  wind  will  not  strike  it  at  any  angle 

441 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

above  the  horizontal.  Chimney  walls  should  be  not  less  than  one  brick  in  width,  and 
they  should  be  lined  with  fire  clay  tile  of  the  size  required  for  the  attached  heating  unit. 
Tile  lining  sizes  are  stated  as  outside  dimensions;  therefore,  their  effective  dimensions 
are  less  by  the  thickness  of  the  wall. 

17  •  Wnat  is  the  purpose  of  a  back  draft  diverter  as  used  on  gas  burning  units? 

Since  the  fuel  is  supplied  under  pressure  independent  of  draft  it  is  necessary  to  free  the 
unit  from  the  variable  chimney  draft  and  to  supply  air  for  combustion  in  direct  propor- 
tion to  the  supply  of  fuel  gas.  The  back  draft  diverter  protects  the  pilot  and  burners 
from  down  drafts. 


442 


Chapter  27 

FUELS  AND  COMBUSTION 

Classification  of  Coaly  Air  for  Combustion,  Draft  Required,  Com- 
bustion of  Anthracite,  Firing  Bituminous  Coal,  Burning  Coke, 
Hand  Firing,  Classification  and  Use  of  Oil,  Classification  and 

Use  of  Gas 

THE  choice  of  fuel  for  heating  is  a  question  of  economy,  cleanliness, 
fuel  availability,  operation  requirements,  and  control.    The  principal 
fuels  to  be  considered  are  coal,  oil,  and  gas. 

COAL 

The  complex  composition  of  coal  makes  it  difficult  to  classify  it  into 
clear-cut  types.  Its  chemical  composition  is  some  indication  but  coals 
having  the  same  chemical  analysis  may  have  distinctly  different  burning 
characteristics.  Users  are  mainly  interested  in  the  available  heat  per 
pound  of  coal,  in  the  handling  and  storing  properties,  and  in  the  burning 
characteristics.  A  description  of  the  relationship  between  the  qualities 
of  coals  and  these  characteristics  requires  considerable  space;  a  treatment 
applicable  to  heating  boilers  is  given  in  U.  S.  Bureau  of  Mines  Bulletin  276. 

A  classification  of  coals  is  given  in  Table  1,  and  a  brief  description  of  the 
kinds  of  fuels  is  given  in  the  following  paragraphs,  but  it  should  be 
recognized  that  there  are  no  distinct  lines  of  demarcation  between  tbf 
kinds,  and  that  they  graduate  into  each  other: 

Anthracite  is  a  clean,  dense,  hard  coal  which  creates  very  little  dust  in  handling.  It 
is  comparatively  hard  to  ignite  but  it  burns  freely  when  well  started.  It  is  non-caking, 
it  burns  uniformly  and  smokelessly  with  a  short  flame,  and  it  requires  little  attention  to 
the  fuel  beds  between  firings.  It  is  capable  of  giving  a  high  efficiency  in  the  common 
types  of  hand-fired  furnaces. 

Semi-anthracite  has  a  higher  volatile  content  than  anthracite,  it  is  not  as  hard  and 
ignites  somewhat  more  easily;  otherwise  its  properties  are  similar  to  those  of  anthracite. 

Semi-bituminous  coal  is  soft  and  friable,  and  fines  and  dust  are  created  by  handling  it. 
It  ignites  somewhat  slowly  and  burns  with  a  medium  length  of  flame.  Its  caking  prop- 
erties increase  as  the  volatile  matter  increases,  but  the  coke  formed  is  relatively  weak. 
Having  only  half  the  volatile  matter  content  of  the  more  abundant  bituminous  coals  it 
can  be  burned  with  less  production  of  smoke,  and  it  is  sometimes  called  smokeless  coal. 

The  term  bituminous  coal  covers  a  large  range  of  coals  and  includes  many  types  having 
distinctly  different  composition,  properties,  and  burning  characteristics.  The  coals  range 
from  the  high-^rade  bituminous  coals  of  the  East  to  the  poorer  coals  of  the  West.  Their 
caking  properties  range  from  coals  which  completely  melt,  to  those  from  which  the 
volatiles  and  tars  are  distilled  without  change  of  form,  so  that  they  are  classed  as  non- 
caking  or  free-burning.  Most  bituminous  coals  are  strong  and  non-friable  enough  to 
permit  of  the  screened  sizes  being  delivered  free  from  fines.  In  general,  they  ignitk 

443 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

easily  and  burn  freely;  the  length  of  flame  varies  with  different  coals,  but  it  is  long.  Much 
smoke  and  soot  are  possible  especially  at  low  rates  of  burning. 

Sub-bituminous  coals  occur  in  the  western  states;  they  are  high  in  moisture  when 
mined  and  tend  to  break  up  as  they  dry  or  when  exposed  to  the  weather;  they  are  liable 
to  ignite  spontaneously  when  piled  or  stored.  They  ignite  easily  and  quickly  and  have  a 
medium  length  flame,  are  non-caking  and  free-burning;  the  lumps  tend  to  break  into 
small  pieces  if  poked;  very  little  smoke  and  soot  are  formed. 

Lignite  is  of  woody  structure,  very  high  in  moisture  as  mined,  and  of  low  heating 
value;  it  is  clean  to  handle.  It  has  a  greater  tendency  than  the  sub-bituminous  coals  to 
disintegrate  as  it  dries,  and  it  also  is  more  liable  to  spontaneous  ignition.  Freshly  mined 
lignite,  because  of  its  high  moisture,  ignites  slowly.  It  is  non-caking.  The  char  left  after 
the  moisture  and  volatile  matter  are  driven  off  burns  very  easily,  like  charcoal.  The 
lumps  tend  to  break  up  in  the  fuel  bed  and  pieces  of  char  falling  into  the  ash  pit  continue 
to  burn.  Very  little  smoke  or  soot  is  formed. 

Coke  is  produced  by  the  distillation  of  the  volatile  matter  from  coal.  The  type  of 
coke  depends  on  the  coal  or  mixture  of  coals  used,  the  temperatures  and  time  of  distil- 
lation and,  to  some  extent,  on  the  type  of  retort  or  oven;  coke  is  also  produced  as  a 
residue  from  the  destructive  distillation  of  oil. 


Legend:  F.C. 


TABLE  1.    CLASSIFICATION  OF  COALS  BY  RANK/ 
Fixed  Carbon.    V.M.  =  Volatile  Matter.    Btu  =  British  thermal  units. 


CLASS 

GROUP 

LIMITS  OP  FIXED  CARBON  on  BTU 
MINERAL-MATTBR-FREE  BASIS 

REQUISITE  PHYSICAL 
PROPERTIES 

1.  Meta-anthracite  

Dry  F.C.,  98  per  cent  or  more  (Dry 

V.M.,  2  per  cent  or  less) 

2.  Anthracite  

Dry  F.C.,  92  per  cent  or  more  and  less 

I.  Anthracite  

than  98  per  cent  (Dry  V.M.,  8  per 

3.  Semi-anthracite  

cent  or  less  and  more  than  2  per  cent) 
Dry  F.C.,  86  per  cent  or  more  and  less 

Non-agglutinating* 

than  92  per  cent  (Dry  V.M.,.  14  per 

cent  or  less  and  more  than  8  per  cent) 

1.  Low  volatile  bituminous  coal  

Dry  F.C.,  77  per  cent  or  more  and  less 

than  86  per  cent  (Dry  V.M.,  23  per 

cent  or  less  and  more  than  14  per 

cent) 

2.  Medium  volatile  bituminous  coal 

Dry  F.C.,  69  per  cent  or  more  and  less 

than  77  per  cent  (Dry  V.M.,  31  per 

cent  or  less  and  more  than  23  per 

II.  Bituminous6™  

cent) 

3.  High  volatile  A  bituminous  coaL 

Dry  P.O.,  leas  than  69  per  cent  (Dry 
V.M.,  more  than  31  per  cent);  and 

moist6  Btu,  14,000*  or  more 

4.  High  volatile  B  bituminous  coal. 

Moist6  Btu,  13,000  or  more  and  less 

than  14,000* 

5.  High  volatile  (7  bituminous  coal.. 

Moist  Btu,  11,000  or  more  and  less 

Either  agglutinating 

than  13,000* 

or  non-weathering* 

1.  Sub-bituminous  A  coaL  

Moist  Btu,  11.000  or  more  and  less 

Both  weathering  and 

than  13,000* 

non-agglutinating 

III.  Sub-bituminous..- 

2.  Sub-bituminous  B  coaL  

Moist  Btu  9500  or  more  and  less 
than  11,  000* 

3.  Sub-bituminous  C  coal  

Moist  Btu  8300  or  more  and   less 

than  9500* 

/ 

1    Lignite 

Moist  Btu  less  than  8300 

Consolidated 

IV.  Lignitic  < 

2.  Brown  coaL  «...  

Moist  Btu  less  than  8300 

Unconsolidated 

fllf  agglutinating,  classify  in  low-volatile  group  of  the  bituminous  class. 

*Moist  Btu  refers  to  coal  containing  its  natural  bed  moisture  but  not  including  visible  water  on  the 
surface  of  the  coal. 

"Pending  the  report  of  the  Subcommittee  on  Origin  and  Composition  and  Methods  of  Analysis,  it  is 
recognized  that  there  may  be  non-caking  varieties  in  each  group  of  the  bituminous  class. 

*Cqals  having  69  per  cent  or  more  fixed  carbon  on  the  dry,  mineral-matter-free  basis  shall  be  classified 
according  to  fixed  carbon,  regardless  of  Btu. 

•There  are  three  varieties  of  coal  in  the  High-volatile  C  bituminous  coal  group,  namely,  Variety  1, 
agglutinating  and  non-weathering;  Variety  2,  agglutinating  and  weathering;  Variety  3,  non-agglutinating 
and  non- weathering. 

/Adapted  from  A.S.T.M.  Standards  on  Coal  and  Coke,  p.  68,  American  Society  for  Testing  Materials, 
Philadelphia,  1934. 


444 


CHAPTER  27 — FUELS  AND  COMBUSTION 


High-temperature  cokes.  Coke  as  usually  available  is  of  the  high-temperature  type, 
and  contains  between  1  and  2  per  cent  volatile  matter.  High-temperature  cokes  are  sub- 
divided into  beehive  coke  of  which  comparatively  little  is  now  sold  for  domestic  use,  by- 
product coke,  which  covers  the  greater  part  of  the  coke  sold,  and  gas-house  coke.  The 
differences  among  these  three  cokes  are  relatively  small ;  their  denseness  and  hardness 
decrease  and  friability  increases  in  the  order  named.  In  general,  the  lighter  and  more 
friable  cokes  ignite  and  burn  the  more  easily. 

Low-temperature  cokes  are  produced  at  low  coking  temperatures,  and  only  a  portion 
of  the  volatile  matter  is  distilled  off.  Cokes  as  made  by  various  processes  under  develop- 
ment have  contained  from  10  to  15  per  cent  volatile  matter.  In  general,  these  cokes 
ignite  and  burn  more  readily  than  high-temperature  cokes.  The  properties  of  various 
low-temperature  cokes  may  differ  more  than  those  of  the  various  high-temperature  cokes 
because  of  the  differences  in  the  quantities  of  volatile  matter  and  because  some  may  be 
light  and  others  briquetted. 

The  sale  of  petroleum  cokes  for  domestic  furnaces  has  been  small  and  is  generally 
confined  to  the  Middle  West.  They  vary  in  the  amount  of  volatile  matter  they  contain, 
but  all  have  the  common  property  of  a  very  low  ash  content,  which  necessitates  the 
use  of  refractory  pieces  to  protect  the  grates  from  being  burned. 

In  order  to  obtain  perfect  combustion  a  definite  amount  of  air  is  re- 
quired for  each  pound  of  fuel  fired.  A  deficiency  of  air  supply  will  result 
in  combustible  products  passing  to  the  stack  unburned.  An  excess  of  air 
absorbs  heat  from  the  products  of  combustion  and  results  in  a  greater  loss 
of  sensible  heat  to  the  stack. 

Total  Air  Required.  The  theoretical  amount  of  air  required  per  pound 
of  fuel  for  perfect  combustion  is  dependent  upon  the  analysis  of  the  fuel ; 

TABLE  2.    POUNDS  OF  AIR  PER  POUND  OF  FUEL  AS  FIRED 


ANTHRACITE 

COKE 

SEMI-BITUMINOUS 

BITUMINOUS 

LIGNITE 

9.6 

11.2 

11.2 

10.3 

6.2 

however,  for  estimating  purposes  the  theoretical  air  required  for  different 
grades  of  fuel  may  roughly  be  taken  from  Table  2.  An  excess  of  about 
50  per  cent  over  the  theoretical  amount  is  considered  good  practice  under 
usual  operating  conditions. 

The  amount  of  excess  air,  based  upon  the  laws  of  combustion,  can  be 
determined  by  its  relation  to  the  percentage  of  COz  (carbon  dioxide)  in 
the  products  of  combustion.  This  relationship  is  shown  by  the  curves 
(Fig.  1)  for  high  and  low  volatile  coals  and  for  coke.  In  hand-fired  fur- 
naces with  long  periods  between  firings  the  combustion  goes  through  a 
cycle  in  each  period  and  the  quantity  of  excess  air  present  varies. 

Secondary  Air.  The  division  of  the  total  into  primary  and  secondary 
air  necessary  to  produce  the  same  rate  of  burning  and  the  same  excess  air 
depends  on  a  number  of  factors  which  include  size  of  fuel,  depth  of  fuel 
bed,  and  diameter  of  fire  pot.  The  ratio  of  the  secondary  to  the  primary 
air  increases  with  decrease  in  the  size  of  the  fuel  pieces,  with  increase  in 
the  depth  of  the  fuel  bed,  and  with  increase  in  the  area  of  the  fire  pot;  the 
ratio  also  increases  with  increase  in  rate  of  burning. 

Size  of  the  fuel  is  a  very  important  factor  in.  fixing  the  quantity  of 
secondary  air  required  for  non-caking  coals.  With  caking  coals  it  is  not 

445 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

- — — * — 

so  important  because  small  pieces  fuse  together  and  form  large  lumps. 
Fortunately  a  smaller  size  fuel  gives  more  resistance  to  air  flow  through 
the  fuel  bed  and  thus  automatically  causes  a  larger  draft  above  the  fuel 
bed,  which  draws  in  more  secondary  air  through  the  same  slot  openings. 
In  spite  of  this,  a  small  size  fuel  requires  a  larger  opening  of  the  door 
slots;  for  a  certain  size  for  each  fuel  no  slot  opening  is  required,  and  for 
larger  sizes  too  much  excess  air  gets  through  the  fuel  bed. 

It  is  impossible  to  establish  a  single  rule  for  the  correct  slot  opening  for 
all  types  and  sizes  of  fuels  and  for  all  rates  of  burning.   Furthermore,  the 


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EXCESS  AIR,  PER  CENT 


FIG.  1.    RELATION  BETWEEN  CO2  AND  EXCESS  AIR  IN  GASES  OF  COMBUSTION 


size  of  slot  opening  is  dependent  on  whether  the  ashpit  damper  is  open 
or  closed.  It  is  better  to  have  too  much  than  too  little  secondary  air ;  the 
opening  is  too  small  if  there  is  a  puff  of  flame  when  the  firing  door  is  opened. 

Fig.  2,  taken  from  the  U.  S.  Bureau  of  Mines  Report  of  Investigations 
No.  2980,  shows  the  relationship  of  the  slot  opening,  for  a  domestic  fur- 
nace, to  the  size  of  coke  and  the  rate  of  burning;  these  openings  are  with 
the  ashpit  damper  wide  open,  and  would  be  less  if  the  available  draft 
permits^of  its  being  partly  closed.  The  same  openings  are  satisfactory  for 
anthracite. 

Bituminous  coals  require  a  large  amount  of  secondary  air  during  the 
period  subsequent  to  a  firing  in  order  to  consume  the  gases  and  to  reduce 
the  smoke.  The  smoke  produced  is  a  good  indicator,  and  that  opening  is 
best  which  reduces  the  smoke  to  a  minimum.  Too  much  secondary  air 
will  cool  the  gases  below  the  ignition  point,  and  prove  harmful  instead  of 
beneficial.  The  following  suggestions  will  be  helpful: 

1.  In  cold  weather,  with  high  combustion  rates,  the  secondary  air  damper  should  be 
half  open  all  the  time. 

2.  In  very  mild  weather,  with  a  very  low  combustion  rate,  the  secondary  air  damper 
should  be  closed  all  the  time. 

,446 


CHAPTER  27 — FUELS  AND  COMBUSTION 


DOOR  SLOTS,  PER  CENT  OPEN 
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SLOTS  AND  DOOR 

5678. 
RATE  OF  BURNIKG,  POUNDS  PER  SQ.  FT.  PER  HR. 

From  U.  S.  Bureau  of  Mines. 

FIG.  2.    RELATIVE  AMOUNT  OF  FIRE  DOOR  SLOT  OPENING  REQUIRED  IN  A  GIVEN 
FURNACE  TO  GIVE  EQUALLY  GOOD  -COMBUSTION  FOR  HIGH  TEMPERATURE 
COKE  OF  VARIOUS  SIZES  WHEN  BURNED  AT  VARIOUS  RATES 

3.  For  temperatures  between  very  mild  and  very  cold,  the  secondary  air  damper 
should  be  in  an  intermediate  position. 

4.  For  ordinary  house  operation,  secondary  air  is  needed  after  each  firing  for  about 
one  hour. 

Draft  Required 

The  draft  required  to  effect  a  given  rate  of  burning  the  fuel  as  measured 
at  the  smokehood  is  dependent  on  the  following  factors : 

1.  Kind  and  size  of  fuel. 

2.  Combustion  rate  per  square  foot  of  grate  area  per  hour. 

3.  Thickness  of  fuel  bed. 

4.  Type  and  amount  of  ash  and  clinker  accumulation. 

5.  Amount  of  excess  air  present  in  the  gases. 

6.  Resistance  offered  by  the  boiler  passes  to  the  flow  of  the  gases. 

7.  Accumulation  of  soot  in  the  passes. 

Insufficient  draft  will  necessitate  additional  manipulation  of  the  fuel 
bed  and  more  frequent  cleanings  to  keep  its  resistance  down.  Insufficient 
draft  also  restricts  the  control  by  adjustment  of  the  dampers. 

The  quantity  of  excess  air  present  has  a  marked  effect  on  the  draft 
required  to  produce  a  given  rate  of  burning,  and  it  is  often  possible  to 
produce  a  higher  rate  by  increasing  the  thickness  of  the  fuel  bed. 

Combustion  of  Anthracite1 

An  anthracite  fire  should  never  be  poked,  as  this  serves  to  bring  ash  to 
the  surface  of  the  fuel  bed  where  it  melts  into  clinker. 
Egg  size  is  suitable  for  large  firepots  (grates  24  in.  and  over)  if  the  fuel 


*See  reports  published  by  The  Anthracite  Institute  Laboratory,  Primoa,  Pennsylvania. 

447 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

can  be  fired  at  least  16  in.  deep.  The  air  spaces  between  the  pieces  of  coal 
are  large,  and  for  best  results  this  coal  should  be  fired  deeply. 

Stove  size  coal  is  the  proper  size  of  anthracite  for  many  boilers  and 
furnaces  used  for  heating  buildings.  It  burns  well  on  grates  at  least  16  in. 
in  diameter  and  12  in.  deep.  The  only  instructions  needed  for  burning 
this  type  of  fuel  are  that  the  grate  should  be  shaken  daily,  the  fire  should 
never  be  poked  or  disturbed,  and  the  fuel  should  be  fired  deeply  and 
uniformly. 

Chestnut  size  coal  is  in  demand  for  firepots  up  to  20  in.  in  diameter,  with 
a. depth  of  from  10  to  15  in. 

Pea  size  coal  is  often  an  economical  fuel  to  burn.  It  is  relatively  low 
in  price.  When  fired  carefully,  pea  coal  can  be  burned  on  standard  grates. 
It  is  well  to  have  a  small  amount  of  a  larger  fuel  on  hand  when  building 
new  fires,  or  when  filling  holes  in  the  fuel  bed.  Care  should  be  taken  to 
shake  the  grates  only  until  the  first  bright  coals  begin  to  fall  through  the 
grates.  The  fuel  bed,  after  a  new  fire  has  been  built,  should  be  increased 
in  thickness  by  the  addition  of  small  charges  until  it  is  at  least  level  with 
the  sill  of  the  fire  door.  This  keeps  a  bed  of  ignited  coal  in  readiness 
against  the  time  when  a  sudden  demand  for  heat  shall  be  made  on  the 
heater. 

Pea  size  coal  requires  a  strong  draft  and  therefore  the  best  results 
generally  will  be  obtained  by  keeping  the  choke  damper  open,  the  cold- 
air  check  closed,  and  by  controlling  the  fire  with  the  air-inlet  damper  only. 
Pea  size  can  also  be  fired  in  layers  with  stove  or  egg  size  anthracite  and 
its  use  in  this  manner  will  reduce  the  fuel  costs  and  attention  required. 

Buckwheat  size  coal  requires  much  the  same  attention  as  pea  size  coal, 
except  that  the  smaller  size  of  the  fuel  makes  it  more  difficult  to  burn  on 
ordinary  grates.  Even  greater  care  must  be  taken  in  shaking  the  grates 
than  with  pea  coal  on  account  of  the  danger  of  the  fuel  falling  through 
the  grate.  A  good  draft  is  required  and  consequently  the  fire  is  best 
controlled  by  the  air-inlet  damper  only.  Where  frequent  attention  can 
be  given  and  where  there  is  not  a  big  heat  demand,  this  fuel  is  frequently 
burned  without  the  aid  of  any  special  equipment. 

In  general  it  will  be  found  more  satisfactory  with  buckwheat  coal  to 
maintain  a  uniform  heat  output  and  consequently  to  keep  the  system 
warm  all  the  time,  rather  than  to  allow  the  system  to  cool  off  at  times 
and  then  to  attempt  to  burn  the  fuel  at  a  high  rate  while  warming  up.  A 
uniform  low  fire  will  minimize  the  clinker  formation  and  keep  the  clinker 
in  an  easily  broken  up  Condition  so  that  it  readily  can  be  shaken  through 
the  grate. 

Forced  draft  and  special  grates  or  retorts  frequently  are  used  with  this 
fuel  for  best  results. 

No.  2  buckwheat  anthracite,  or  rice  size,  is  used  only  with  forced  draft 
equipment  on  mechanical  stokers.  No.  3  buckwheat  anthracite,  or  barley, 
has  no  application  in  domestic  heating. 

Firing  Bituminous  Coal 

Bituminous  coal  should  never  be  fired  over  the  entire  fuel  bed  at  one 
time.  A  portion  of  the  glowing  fuel  should  always  be  left  exposed  to 
ignite  the  gases  leaving  the  fresh  charge. 

448 


CHAPTER  27 — FUELS  AND  COMBUSTION 


Air  should  be  admitted  over  the  fire  through  a  special  secondary  air 
device,  or  through  a  slide  in  the  fire  door  or  by  opening  the  fire  door 
slightly.  If  the  quantity  of  air  admitted  is  too  great  the  gases  will  be 
cooled  below  the  ignition  temperature  and  will  fail  to  burn.  The  fireman 
can  judge  the  quantity  of  air  to  admit  by  noting  when  the  air  supplied 
is  just  sufficient  to  make  the  gases  burn  rapidly  and  smokelessly  above  the 
fuel  bed. 

The  red  fuel  in  the  firebox,  before  firing,  excepting  only  a  shallow  layer 
of  coke  on  the  grate,  should  be  pushed  to  one  side  or  forward  or  back- 
ward to  form  a  hollow  in  which  to  throw  the  fresh  fuel.  Some  manu- 
facturers recommend  that  all  red  fuel  be  pushed  to  the  rear  of  the  firebox 
and  that  the  fresh  fuel  be  fired  directly  on  the  grate  and  allowed  to  ignite 
from  the  top.  The  object  of  this  is  to  reduce  the  early  rapid  distillation 
of  gases  and  to  reduce  the  quantity  of  secondary  air  required  for  smoke- 
less combustion. 

It  is  well  to  have  the  bright  fuel  in  the  firebox  so  placed  that  the  gases 
from  the  freshly  fired  fuel,  mixed  with  the  air  over  the  fuel  bed,  pass 
over  the  bed  of  bright  fuel  on  the  way  to  the  flues.  The  bed  of  bright 
fuel  then  supplies  the  heat  to  raise  the  mixture  of  air  and  gas  to  the 
ignition  temperature,  thereby  causing  the  gaseous  matter  to  burn  and 
preventing  the  formation  of  smoke. 

The  fuel  bed  should  be  carried  as  deep  as  the  size  of  fuel  and  the 
available  draft  permit,  in  order  to  have  as  much  coked  fuel  as  possible 
for  pushing  to  the  rear  of  the  firebox  at  the  time  of  firing.  A  deep  fuel 
bed  allows  the  longest  firing  intervals. 

If  the  coal  is  of  the  caking  kind  the  fresh  charge  will  fuse  into  one 
solid  mass  which  can  be  broken  up  with  the  stoking  bar  and  leveled  from 
20  min  to  one  hour  after  firing,  depending  on  the  temperature  of  the 
firebox.  Care  should  be  exercised  when  stoking  not  to  bring  the  bar  up 
to  the  surface  of  the  fuel  as  this  will  tend  to  bring  ash  into  the  high 
temperature  zone  at  the  top  of  the  fire,  where  it  will  melt  and  form 
clinker.  The  stoking  bar  should  be  kept  as  near  the  grate  as  possible 
and  should  be  raised  only  enough  to  break  up  the  fuel.  With  fuels  requir- 
ing stoking  it  may  not  be  necessary  to  shake  the  grates,  as  the  ash  is 
usually  dislodged  during  stoking. 

The  output  obtained  from  any  heater  with  bituminous  coal  will  usually 
exceed  that  obtainable  with  anthracite,  since  soft  coal  burns  more  rapidly 
than  hard  coal  and  with  less  draft.  Soft  coal,  however,  will  require 
frequent  attention  to  the  fuel  bed,  because  it  burns  unevenly,  even 
though  the  fuel  bed  may  be  level,  forming  holes  in  the  fire  which  admit 
too  much  air,  chilling  the  gases  over  the  fuel  bed  and  reducing  the 
available  draft. 

Semi-bituminous  coal  is  fired  as  bituminous  coal,  and  because  of  its 
caking  characteristics  it  requires  practically  the  same  attention.  The 
Pocahontas  Operators  Association  recommends  the  central  cone  method  of 
firing,  in  which  the  coal  is  heaped  on  to  the  center  of  the  bed  forming  a 
cone  the  top  of  which  should  be  level  with  the  middle  of  the  firing  door. 
This  allows  the  larger  lumps  to  fall  to  the  sides,  and  the  fines  to  remain  in 
the  center  and  be  coked.  The  poking  should  be  limited  to  breaking  down 
the  coke  without  stirring,  and  to  gently  rocking  the  grates.  It  is  recom- 

449 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

mended  that  the  slides  in  the  firing  door  be  kept  closed,  as  the  thinner  fuel 
bed  around  the  sides  allows  enough  air  to  get  through. 

Burning  Coke 

Coke  is  a  very  desirable  fuel  and  usually  will  give  satisfaction  as  soon 
as  the  user  learns  how  to  control  the  fire.  Coke  ignites  and  burns  very 
rapidly  with  less 'draft  than  anthracite  coal.  In  order  to  control  the  air 
admitted  to  the  fuel  it  is  very  important  that  all  openings  or  leaks  into 
the  ashpit  be  closed  tightly.  A  coke  fire  responds  more  rapidly  than  an 
anthracite  fire  to  the  opening  of  the  dampers.  This  is  an  advantage  in 
warming  up  the  system,  but  it  also  makes  it  necessary  to  watch  the 
dampers  more  closely  in  order  to  prevent  the  fire  from  burning  too  rapidly. 
A  deep  fuel  bed  always  should  be  maintained  when  burning  coke.  The 
grates  should  be  shaken  only  slightly  in  mild  weather  and  should  be 
shaken  only  until  the  first  red  particles  drop  from  the  grates  in  cold 
weather.  Since  coke  weighs  only  about  half  as  much  as  anthracite  per 
cubic  foot  only  about  half  as  much  can  be  put  in  the  firepot,  so  it  will  be 
necessary  to  fire  oftener.  The  best  size  of  coke  for  general  use,  for  small 
firepots  where  the  fuel  depth  is  not  over  20  in.,  is  that  which  passes  over 
a  1  in.  screen  and  through  a  \Y%  in.  screen.  For  large  firepots  where  the 
fuel  can  be  fired  over  20  in.  deep,  coke  which  passes  over  a  1  in.  screen  and 
through  a  3  in.  screen  can  be  used,  but  a  coke  of  uniform  size  is  always 
more  satisfactory.  Large  sizes  of  coke  should  be  either  mixed  with  fine 
sizes  or  broken  up  before  using. 

Dustless  Coal 

The  practice  of  treating  the  more  friable  coals  to  allay  the  dust  they 
create  is  increasing.  The  coal  is  sprayed  with  a  solution  of  calcium 
chloride  or  a  mixture  of  calcium  and  magnesium  chlorides.  Both  these 
salts  are  very  hygroscopic  and  their  moisture  under  normal  atmospheric 
conditions  keeps  the  surface  of  the  coal  damp,  thus  reducing  the  dust 
during  delivery  and  in  the  cellar,  and  obviating  the  necessity  of  sprinkling 
the  coal  in  the  bin. 

The  coal  is  sometimes  treated  at  the  mine,  but  more  usually  by  the 
local  distributor  just  before  delivery.  The  solution  is  sprayed  under  high 
pressure,  using  from  2  to  4  gal  or  from  5  to  10  Ib  of  the  salt  per  ton  of 
coal,  depending  on  its  friability  and  size. 

Pulverized  Coal 

Installations  of  pulverized  coal  burning  plants  in  heating  boilers  are  of 
the  unit  type,  in  which  the  pulverized  coal  is  delivered  into  the  furnace 
immediately  after  grinding,  together  with  the  proper  amount  of  preheated 
air.  With  this  apparatus,  where  the  necessary  furnace  volume  is  ob- 
tainable, high  efficiencies  can  be  obtained. 

A  150-hp  boiler  has  generally  been  considered  the  smallest  size  for 
which  pulverized  fuel  is  feasible.  Complications  are  introduced  if  an 
installation  with  a  single  boiler  has  to  take  care  of  very  light  loads. 

Hand  Firing 

Hand  firing  is  the  oldest  and  the  most  widely  used  method  of  burning 
coal  for  heating  purposes.  To  keep  the  fuel  bed  in  proper  condition  where 
hand  firing  is  used,  the  following  general  rules  should  be  observed : 

450 


CHAPTER  27 — FUELS  AND  -COMBUSTION 


1.  Remove  ash  from  fuel  bed  by  shaking  the  grates  whenever  fresh  fuel  is  fired.    This 
removes  ashjrom  the  fire,  enables  the  air  to  reach  the  fuel,  and  does  away  with  the  for- 
mation of  clinker  which  is  melted  ash. 

2.  Supply  the  boiler  with  a  deep  bed  of  fuel.    Nothing  is  gained  by  attempting  to 
fire  a  small  amount  of  fuel.    A  deep  bed  of  fuel  secures  the  most  economical  results. 

3.  Remove  ash  from  ashpit  at  least  once  daily.    Never  allow  ash  to  accumulate  up 
to  the  grates.    If  the _ ash  prevents  the  air  from  passing  through,  the  grate  bars  will 
burn  out  and  much  clinker  trouble  .will  be  experienced. 

The  principal  requirements  for  a  hand- fired  furnace  are  that  it  shall  have 
enough  grate  area  and  combustion  space.  The  amount  of  grate  area 
required  is  dependent  upon  the  desired  combustion  rate. 

The  furnace  volume  is  influenced  by  the  kind  of  coal  used.  Bituminous 
coals,  on  account  of  their  long-flaming  characteristic,  require  more  space 
in  which  to  burn  the  gases  of  combustion  completely  than  do  the  coals 
low  in  volatile  matter.  For  burning  high  volatile  coals  provision  should 
be  made  for  mixing  the  combustible  gases  thoroughly  so  that  com- 
bustion is  complete  before  the  gases  come  in  contact  with  the  relatively 
cool  heating  surfaces.  An  abrupt  change  in  the  direction  of  flow  tends  to 
mix  the  gases  of  combustion  more  thoroughly. 

OIL 

Uniform  oil  specifications  were  prepared  in  1929  by  the  American  Oil 
Burner  Association,  in  cooperation  with  the  American  Petroleum  Institute, 
the  U.  S.  Bureau  of  Standards,  the  American  Society  for  Testing  Materials 
and  other  interested  organizations.  Oil  fuels  were  classified  into  six 
groups,  as  indicated  by  Table  3.  When  these  specifications  were  prepared, 
it  was  generally  accepted  that  the  first  three  grades  were  adapted  to 
domestic  use,  while  the  last  three  were  suitable  only  for  commercial  and 
industrial  burners.  Today  domestic  installations  are  using  No.  4  of  the 
so-called  heavy-oil  group,  due  principally  to  the  fact  that  No.  4  oils  in 
general  are  being  offered  of  better  grade  and  adaptability  than  those  called 
for  in  the  commercial  specifications. 

Since  the  specifications  as  originally  drawn  provide  for  maximum  limits 
only  for  the  several  grades,^  this  differentiation  has  not  proved  stable. 
Realizing  how  unsatisfactory  it  is  to  have  specifications  which  permit  the 
substitution  of  one  grade  for  another,  the  U.  S.  Bureau  of  Standards  in 
cooperation  with  the  American  Society  for  Testing  Materials  is  figuring 
on  a  new  set  of  specifications  providing  for  definite  limits  for  each  grade. 
When  these  specifications  are  adopted,  it  is  expected  that  the  National 
Board  of  Fire  Underwriters  will  retest  all  burners  using  oils  of  the  maximum 
specifications  for  the  grade  so  that  if  a  burner  is  approved  for  a  certain 
grade  it  will  burn  any  oil  meeting  the  specifications  for  that  particular 
grade. 

Several  burners  adapted  to  industrial  use  have  recently  been  listed  for 
automatic  operation  with  No.  5  oil.  Usually  oils  No.  5  or  6  require 
preheating  for  proper  operation,  but  where  conditions  are  favorable,  No. 
5  can  be  used  without  the  equipment  that  this  entails. 

There  are  two  reasons  for  the  trend  to  lower  grades  of  oil.  While  the 
lighter  oils  contain  slightly  more  heat  units  per  pound,  the  weight  per 

451 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


TABLE  3.    COMMERCIAL  STANDARD  FUEL  OIL  SPECIFICATIONS0 
A .   Detailed  Requirements  for  Domestic  Fuel  Oils 


GRADE  OF  OIL 

APPROX. 
BTU 

PER 

GAL.& 

FLASH   POINT 

WATER 

AND 

SEDIMENT, 
MAXIMUM 

POUR 
POINT,* 

MAXIMUM 

DISTILLATION 
TEST 

VISCOSITY 
MAXIMUM 

Min. 

Max. 

No.  1 
Domestic 
Fuel  Oil 
A  light  distillate 
oil    for    use    in 
burners  requir- 
ing a  high  grade 
fuel. 

139.000 

110  F 

or  legal 

165  F 

0.05% 

15  F 

10%  point, 
maximum 

420  F 

End  point, 
maximum 

600  F 

No.  2 
Domestic 
Fuel  Oil 
A  medium  distil- 
late oil  for  use 
in   burners   re- 
quiring a   high 
grade  fuel. 

141.000 

125  F 

or  legal 

190  F 

0.05% 

15  F 

10%  point, 
maximum 

440  F 

90%  point, 
maximum 

620  F 

No.  3 
Domestic 
Fuel  Oil 
A    distillate    fuel 
oil   for   use   in 
burners  where  a 
low  viscosity  oil 
is  required. 

143,400 

150  F 
or  legal 

200  F 

0.1% 

15  F 

10%  point, 
maximum 

460  F 

90%  point, 
maximum 

675  F 

Saybolt 
Universal 
at  100  F 
55  seconds 

B.   Detailed  Requirements  for  Industrial  Fuel  Oils 


GRADE  OF  OIL 

APPROX. 

BTU 

PER 

GAL.* 

FLASH  POINT, 
MIN.    MAX. 

WATER 

AND 

SEDIMENT, 
MAXIMUM 

POUR 
POINT,* 
MAXIMUM 

VISCOSITY, 
MAXIMUM 

No.  4. 

Industrial  Fuel  Oil 

An  oil  known  to  the  trade  as  a  light  fuel 
oil  for  use  in  burners  where  a  low  vis- 
cosity industrial  fuel  oil  is  required. 

144,500 

150  F.    See 

Note* 

1.0% 

See 
Note- 

Saybolt 
Universal 
at  100  F 
125  seconds 

No.  5 
Industrial  Fuel  Oil 

Same  as  Federal  Specifications   Board 
specification  for  bunker  oil  "B"  for 
burners  adapted  to  the  use  of  indus- 
trial fuel  oil  of  medium  viscosity. 

146,000 

150  F 

1.0% 

Saybolt 
Furol 
at  122  F 
100  seconds 

No.  6 
Industrial  Fuel  Oil 

Same  as  Federal  Specifications  Board 
specification  for  bunker  oil  "C"  for 
burners  adapted  to  oil  of  high  viscosity. 

150,000 

150  F 

.Water 
sediment 

1.75% 
0.25% 

Saybolt 
Furol 
at  122  F 
300  seconds 

°Adapted  from  "Fuel  Oils,"  p.  2,  U,  S.  Department  of  Commerce,  Bureau  of  Standards,  Commercial 
Standard  CS1Z-SS,  Washington,  1933. 

^Government  specifications  do  not  give  Btu  per  gallon,  but  they  are  noted  here  for  information  only. 

'Lower  or  higher  pour  points  may  be  specified  whenever  required  by  conditions  of  storage  and  use. 
However,  these  specifications  shall  not  require  a  pour  point  less  than  0  F  under  any  conditions. 

^Whenever  required,  as  for  example  in  burners  with  automatic  ignition,  a  maximum  flash  point  may 
be  specified.  However,  these  specifications  shall  not  require  a  flash  point  less  than  250  F  under  any 
conditions. 

'Pour  point  may  be  specified  whenever  required  by  conditions  of  storage  and  use.  However,  these 
specifications  shall  n6t  require  a  pour  point  less  than  15  F  under  any  conditions. 


452 


CHAPTER  27 — FUELS  AND  COMBUSTION 


gallon  increases  more  rapidly  than  the  decrease  in  heat  units  per  pound, 
and  oil  is  bought  by  the  gallon.  As  a  consequence,  while  a  No.  1  oil  may 
contain  139,000  Btu  per  gallon,  oil  No.  5  may  test  146,000  Btu  per  gallon, 
or  6  per  cent  more.  Usually  there  is  a  differential  of  3  i  to  4^  between  the 
No.  1  and  No.  5  oils,  so  that  the  economy  of  buying  the  heavier  fuels  is 
apparent;  there  remains  the  economic  utilization  of  the  heat  content  of 
the  heavier  oils. 

The  cost  of  oil  fuel  is  dependent  also  upon  the  amount  that  can  be 
delivered  at  one  time,  and  the  method  of  delivery.  Common  practice  has 
split  the  tank  of  the  truck  delivering  oils  for  domestic  use  into  compart- 
ments of  150  to  500-gal  capacity,  and  these  unit  dumps  are  made  the  basis 
of  price.  Where  a  truck  can  be  connected  to  a  storage-tank  fill  and 
quickly  discharge  its  oil  by  pump,  the  price  obviously  can  be  less  than 
where  a  smaller  quantity  must  be  drawn  off  in  5-gal  cans  and  poured. 
For  similar  reasons  an  installation  that  can  be  supplied  from  a  tank  car 
on  a  siding  provides  for  a  lower  unit  fuel  cost  than  one  where  the  oil 
must  be  trucked,  even  in  the  large  trucks  holding  2,000  gal  or  more  that 
are  used  for  distributing  the  heavier  oils. 

GAS 

Gas  is  broadly  classified  as  being  either  natural  or  manufactured. 
Natural  gas  is  a  mechanical  mixture  of  several  combustible  and  inert 
gases  rather  than  a  chemical  compound.  Manufactured  gas  as  dis- 
tributed is  usually  a  combination  of  certain  proportions  of  gases  produced 
by  two  or  more  processes,  and  is  often  designated  as  city  gas. 

When  gas  is  burned  a  large  amount  of  water  vapor  is  produced  as  one 
of  the  products  of  combustion.  This  ordinarily  escapes  up  the  chimney, 
carrying  away  with  it  a  certain  amount  of  heat.  However,  when  the  heat 
value  of  gas  is  determined  in  an  ordinary  calorimeter,  this  water  vapor 
is  condensed  and  the  latent  heat  of  vaporization  that  is  given  up  during 
the  condensation  is  reported  as  a  portion  of  the  heat  value  of  the  gas. 
The  heat  value  so  determined  is  termed  the  gross  or  higher  heat  value  and 
this  is  what  is  ordinarily  meant  when  the  heat  value  of  gas  is  specified. 
The  heat  that  is  reclaimed  by  the  condensation  of  the  water  vapor 
amounts  to  about  10  per  cent  of  the  total  heat  value.  It  is  impractical 
to  utilize  the  entire  higher  heat  value  of  the  gas  in  any  house^heating 
appliance,  because  to  do  so  it  would  be  necessary  to  cool  the  products  of 
combustion  down  below  their  dew  point,  which  is  ordinarily  in  the 
neighborhood  of  130  F. 

Natural  gas  is  the  richest  of  the  gases  and  contains  from  80  to  95 
per  cent  methane,  with  small  percentages  of  the  other  combustible 
hydrocarbons.  In  addition,  it  contains  from  0.5  to  5.0  per  cent  of  COz, 
and  from  1  to  12  or  14  per  cent  of  nitrogen.  The  heat  value  varies  from 
700  to  1,500  Btu  per  cubic  foot,  the  majority  of  natural  gases  averaging 
about  1,000  Btu  per  cubic  foot.  Table  4  shows  typical  values  for  the 
three  main  oil  fields,  although  values  from  any  one  field  vary  materially. 

Table  4  also  gives  the  calorific  values  of  the  more  common  types  of 
manufactured  gas.  Most  states  have  legislation  which  control^  the  distri- 
bution of  gas  and  fixes  a  minimum  limit  to  its  heat  content.  The  gross 

453 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


or  higher  calorific  value  usually  ranges  between  520  and  545  Btu  per  cubic 
foot,  with  an  average  of  535.  A  given  heat  value  may  be  maintained  and 
yet  leave  considerable  latitude  in  the  composition  of  the  gas  so  that  as 
distributed  the  composition  is  not  necessarily  the  same  in  different  dis- 
tricts, nor  at  successive  times  in  the  same  district.  There  are  limits  to  the 
variation  allowable,  because  the  specific  gravity  of  the  gas  depends  on  its 

TABLE  4.   REPRESENTATIVE  PROPERTIES  OF  GASEOUS  FUELS, 
BASED  ON  GAS  AT  60  F  AND  30  IN.  HG. 


GAS 

BTU  PER  Cu  FT 

SPECIFIC 
GRAVITY, 
AIR  = 
1.00 

Am  REQUIRED 
FOR  COMBUS- 
TION, 
(Cu  FT) 

PRODUCTS  OP  COMBUSTION 

THEORETICAL 
FLAME   TEM- 
PERATURE, 
(DEG  FAHR) 

High 
(Gross) 

Low 

(Net) 

Cubic  Feet 

ULTI- 
MATE 
CO* 
Dry 
Basis 

CQt 

#20 

Total 
with 

N* 

Natural  gas  — 
Mid-Conti- 
nental 

967 

873 

0.57 

9.17 

0.97 

1.92 

10.2 

11.7 

3580 

Natural  gas  — 
Ohio 

1130 

1025 

0.65 

10.70 

1.17 

2.16 

11.8 

12.1 

3600 

Natural  gas  — 
Pennsylvania 

1232 

1120 

0.71 

11.70 

1.30 

2.29 

12.9 

12.3 

3620 

Retort  coal  gas 

575 

510 

0.42 

5.00 

0.50 

1.21 

5.7 

11.2 

3665 

Coke  oven  gas 

588 

521 

0.42 

5.19 

0.51 

1.25 

5.9 

11.0 

3660 

Carburet  ted 
water  gas 

536 

496 

0.65 

4.37 

0.74 

0.75 

5.0 

17.2 

3815 

Blue  water  gas 

308 

281 

0.53 

2.26 

0.46 

0.51 

2.8 

22.3 

3800 

Anthracite  pro- 
ducer gas 

134 

124 

0.85 

1.05 

0.33 

0.19 

1.9 

19.0 

3000 

Bituminous 
producer  gas 

150 

140 

0.86 

1.24 

0.35 

0.19 

2.0 

19.0 

3160 

Oil  gas 

575 

510 

0.35 

4.91 

0.47 

1.21 

5.6 

10.7 

3725 

composition,  and  too  great  a  change  in  the  specific  gravity  necessitates  a 
change  in  the  adjustment  of  the  burners  of  small  appliances. 

Table  4  shows  that  a  large  proportion  of  the  products  of  combustion 
when  gas  is  burned  may  consist  of  water  vapor,  and  that  the  greater  the 
proportion  of  water  vapor,  the  lower  the  maximum  attainable  CO*  by  gas 
analysis.  The  table  also  shows  that  a  low  calorific  value  does  not  neces- 
sarily mean  a  low  flame  temperature  since,  for  example,  natural  gas  has  a 
theoretical  flame  temperature  of  3600  F  and  blue  water  gas  of  3800  F, 
although  it  has  a  calorific  value  less  than  one  third  that  of  natural  gas. 

The  quantity  of  air  given  in  Table  4  is  that  required  for  theoretical 
combustion,  but  with  a  properly  designed  and  installed  burner  the  excess 
air,  can  be  kept  low.  The  division  of  the  air  into  primary  and  secondary- 
is  a  matter  of  burner  design  and  the  pressure  of  gas  available,  and  also  of 
the  type  of  flame  desired. 


'  CHAPTER  27 — FUELS  AND  COMBUSTION 


PROBLEMS  IN  PRACTICE 

1  •  Name  several  important  properties  of  coal  from  a  utilization  standpoint. 

a.  Caking  tendency,  whether  none,  weak,  or  strong. 

b.  Quantity  of  volatile  matter. 

c.  Friability. 

d.  Fusibility  of  the  ash. 

2  •  What  are  the  main  data  commonly  available  that  fix  the  qualities  of  coal, 
and  do  these  tell  the  whole  story? 

a.  Calorific  value,  Btu  per  pound. 

b.  Proximate  analysis  giving  percentages  of  moisture,  volatile  matter,  fixed  carbon,  ash, 
and  sulphur. 

c.  Temperature  at  which  the  ash  softens. 

d.  Screen  sizes. 

Other  important  qualities  not  usually  given  are  the  friability  of  the  coal,  its  caking 
tendency,  and  the  qualities  of  the  volatile  matter.  The  percentage  of  ash  and  its  fusion 
temperature  do  not  tell  how  the  ash  is  distributed  or  how  much  of  it  is  less  fusible  lumps 
of  slate  or  shale. 

3  •  Are  there  available  complete  and  sufficient  data  on  gas  and  oils  to  fix  their 
burning  properties  and  furnace  requirements? 

Yes.  Because  gas  and  oils  are  of  simple  and  uniform  composition,  data  are  available  to 
fix  their  burning  properties  and  furnace  requirements,  but  the  ability  to  control  their 
combustion  is  somewhat  less  determinable. 

4  •  What  effect  does  moisture  in  fuels  have  on  then*  efficiency? 

With  any  solid  fuel,  latent  and  sensible  heat  are  lost  at  the  stack  when  moisture  is  dried 
out  of  the  fuel  in  burning,  and  when  its  hydrogen  is  burned.  Therefore,  such  fuels  as 
sub-bituminous  coal  and  lignite,  which  are  high  in  moisture  content,  have  a  low  efficiency. 
However,  these  efficiencies  may  be  improved  if  the  stack  gases  are  cooled  to  room  tem- 
perature, by  heating  the  feed  water,  for  example. 

5  •  What  are  the  advantages  of  a  sized  fuel  for  heating  furnaces? 

Because  a  sized  fuel  encourages  a  more  uniform  flow  of  air  through  the  bed,  the  burning 
will  be  more  uniform,  and  the  bed  will  be  less  liable  to  develop  holes  and  will  require  less 
attention.  Uniformity  of  fuel  size  is  more  desirable  as  the  area  of  the  bed  becomes 
smaller;  it  is  less  important  with  fuels  that  cake,  but  with  sized  fuels  the  caking  will  be 
more  uniform  and  the  air  flow  through  the  bed  will  be  steadier.  In  addition,  ash  and 
pieces  of  slate  are  less  likely  to  be  segregated  and  to  form  lumps  of  clinker. 

6  •  Does  the  size  of  a  fuel  affect  the  quantity  of  air  required  to  burn  it  at  a 
given  rate? 

The  total  air  required  to  give  the  same  gas  analysis  at  the  stack  is  independent  of  the 
size  of  the  fuel  burned,  but  for  non-caking  fuels  the  ratio  of  the  air  passing  through  the 
fuel  bed  to  the  total  air  entering  the  burner  base  decreases,  for  the  same  thickness  of  bed, 
as  the  size  of  the  fuel  becomes  smaller;  this  decrease  is  very  rapid  for  sizes  less  than  one 
inch.  For  coals  that  cake,  this  ratio  will  depend  on  the  way  the  caked  bed  is  broken  up 
and  on  the  size  of  the  resulting  pieces. 

7  •  Is  the  volatile  matter  which  is  given  off  when  coals  are  burned  of  the  same 
nature  in  all  coals? 

No.  The  products  given  off  by  coals  when  they  are  heated  differ  materially  in  the  ratios 
by  weight  of  the  gases  to  the  oils  and  tars.  No  heavy  oils  or  tars  are  given  off  by  anthra- 
cite, and  very  small  quantities  are  given  off  by  semi-anthracite.  As  the  volatile  matter 
in  the  coal  increases  to  as  much  as  40  per  cent  of  ash-free  and  moisture-free  coal,  in- 

455 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

creasing  amounts  of  oils  and  tars  are  given  up.  For  coals  of  higher  volatile  content,  the 
relative  quantity  of  oils  and  tars  decreases,  so  it  is  low  in  the  sub-bituminous  coals  and 
in  lignite. 

8  •  Is  smoke  a  primary  product  in  the  burning  of  fuels? 

Visible  smoke  may  include  very  small  particles  of  carbon,  oil,  tar,  water  (condensed 
steam),  and  ash.  Of  these,  the  oils,  tars,  and  ash  are  mainly  primary  products,  and  the 
water  is  partly  primary.  The  carbon,  which  usually  comprises  the  greater  part  of  the 
smoke,  results  from  the  breaking  up  by  heat  of  oils,  tars,  and  such  gases  as  methane,  so 
it  may  be  considered  a  secondary  product. 

9  •  Is  the  sulphur  in  coals  detrimental  to  combustion? 

Not  so  far  as  is  known,  but  its  complete  combustion  gives  only  25  per  cent  as  much  heat 
as  is  given  by  the  same  weight  of  carbon.  Sulphur  is  undesirable  because  it  causes  cor- 
rosion of  flues  and  stacks,  and  also  because  its  gases  pollute  the  atmosphere,  and  damage 
buildings  and  vegetation. 

10  •  Can  any  one  fuel  be  said  to  be  the  best'fuel? 

The  term  best  can  be  applied  to  a  fuel  only  after  consideration  of  the  cost  factor  of  the 
fuel  and  the  equipment  necessary  for  its  use.  Gas  seems  to  be  the  most  convenient  fuel 
because  of  its  uniformity,  and  the  ease  with  which  it  may  be  controlled  and  its  burning 
made  fully  automatic. 


456 


Chapter  28 

AUTOMATIC  FUEL  BURNING 
EQUIPMENT 

Stokers,  Residential  Stokers,  Apartment  House  Stokers,  Com- 
mercial Stokers,  Domestic  Oil  Burners,  Commercial  Oil  Burners, 
Gas-Fired  Appliances,  Gas  Boilers,  Warm  Air  Furnaces,  Space 
Heaters,  Conversion  Burners,  Gas  Appliances 

A  UTOMATIC,  mechanical  equipment  for  the  efficient  combustion 
-Z~jL  of  coal,  oil,  and  gas  is  considered  in  this  chapter. 

MECHANICAL  STOKERS 

Assuming  the  same  intelligence  in  handling  the  fire,  coal  can  be  burned 
more  efficiently  on  a  mechanical  stoker  than  on  any  kind  of  hand-fired 
grate.  This  does  not  necessarily  mean  that  a  stoker  installation  may  be 
more  economical,  because  the  amount  of  coal  burned  may  be  so  small  or  the 
cost  of  the  installation  so  high  that  the  savings  with  stokers  may  .not  be 
sufficient  to  pay  for  the  investment.  The  operation  of  burning  coal  in- 
volves uniformity  in  stoking,  proper  distribution  over  the  fuel  bed, 
admission  of  air  as  required  to  all  parts  of  the  fuel  bed,  and  disposal  of  the 
ash.  The  handling  of  the  volatile  gas  is  largely  a  matter  of  furnace  design 
but  since  this  gas  forms  a  considerable  portion  of  the  heating  value  of  the 
coal,  it  may  also  be  said  that  the  proper  handling  of  this  gas  is  a  function 
of  firing.  All  mechanical  stokers  must  provide  means  of  taking  care  of 
these  several  functions  in  order  fully  to  serve  their  purpose. 

Stokers  may  be  divided  into  four  types  according  to  their  construction 
and  operation,  namely,  (1)  overfeed  flat  grate,  (2)  overfeed  inclined  grate, 
(3)  underfeed  side  cleaning  type,  and  (4)  underfeed  rear  cleaning  type. 
They  may  also  be  classified  according  to  their  uses.  The  following 
classification  has  been  adopted  by  the  U.  S.  Department  of  Commerce: 

Class  1.  Residential  (Capacity  less  than  100  Ib  coal  per  hour). 

Class  2.  Apartment  houses  and  small  commercial  heating  jobs  (Capacity  100  to  200  Ib 
coal  per  hour). 

Class  3.  General  commerical  heating  and  small  high  pressure  steam  plants  (Capacity 
200  to  300  Ib  coal  per  hour). 

Class  4.  Large  commercial  and  high  pressure  steam  plants  (Capacity  over  300  Ib 
per  hour). 

Overfeed  Flat  Grate  Stokers 

This  type  is  represented  by  the  various  chain  grate  stokers.  These 
stokers  receive  fuel  at  the  front  of  the  grate  in  a  layer  of  uniform  thickness 
and  move  it  back  horizontally  to  the  rear  of  the  furnace.  Air  is  supplied 

457 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

under  the  moving  grate  to  carry  on  combustion  at  a  sufficient  rate  to  com- 
plete the  burning  of  the  coal  near  the  rear  of  the  furnace.  The  ash  is 
carried  over  the  back  end  of  the  stoker  into  an  ashpit  beneath.  This  type 
of  stoker  is  suitable  for  small  sizes  of  anthracite  or  coke  breeze  and  also 
for  bituminous  coals,  the  clinker  forming  characteristics  of  which  make  it 
desirable  to  burn  the  fuel  without  disturbing  it.  This  type  of  stoker 
invariably  requires  the  use  of  an  arch  over  the  front  of  the  stoker  to 
maintain  ignition  of  the  incoming  fuel  and  to  maintain  the  volatile  gas  at 
a  temperature  suitable  for  combustion.  Frequently,  a  rear  combustion 
arch  is  required  to  maintain  ignition  until  the  fuel  is  fully  consumed. 

Overfeed  Inclined  Crate  Stokers 

In  general  the  combustion  principle  is  similar  to  the  flat  grate  stoker, 
but  this  stoker  is  provided  with  rocking  grates  set  on  an  incline  to  advance 
the  fuel  during  combustion.  Also  this  type  is  provided  with  an  ash  plate 
where  ash  is  accumulated  and  from  which  it  is  dumped  periodically. 
This  type  of  stoker  is  suitable  for  all  types  of  coking  fuels  but  preferably 
for  those  of  low  volatile  content.  Its  grate  action  has  the  tendency  to 
keep  the  fuel  bed  well  broken  up  thereby  allowing  for  free  passage  of  air. 
Because  of  its  agitating  effect  on  the  fuel  it  is  not  so  desirable  for  badly 
clinkering  coals.  Furthermore,  it  should  usually  be  provided  with  a 
front  arch  to  care  for  the  volatile  gas. 

Underfeed  Side  Cleaning  Stokers 

In  this  type,  the  fuel  is  fed  in  at  the  front  of  the  furnace  to  one  or  more 
retorts,  is  advanced  away  from  the  retort  as  combustion  progresses,  while 
finally  the  ash  is  disposed  of  at  the  sides.  This  type  of  stoker  is  suitable 
for  all  coking  coals  while  in  the  smaller  sizes  it  is  suitable  for  small  sizes  of 
anthracites.  In  this  type  of  stoker  the  fuel  is  delivered  to  a  retort  beneath 
the  fire  and  is  raised  into  the  fire.  During  this  process  the  volatile  gas  is 
released,  is  mixed  with  air,  and  passes  through  the  fire  where  it  is  burned. 
The  ash  may  be  continuously  discharged  as  in  the  small  stoker  or  may  b.e 
accumulated  on  a  dump  plate  and  periodically  discharged.  This  stoker 
requires  no  arch  as  it  automatically  provides  for  the  combustion  of  the 
volatile  gas. 

Underfeed  Rear  Cleaning  Stokers 

This  type  carries  on  combustion  in  much  the  same  manner  as  the  side 
cleaning  type,  but  consists  of  several  retorts  placed  side  by  side  and 
filling  up  the  furnace  width,  while  the  ash  disposal  is  at  the  rear.  In 
principle,  its  operation  is  the  same  as  the  side  cleaning  underfeed. 

Class  1  Stokers,  Residential 

A  common  type  of  stoker  in  this  class  consists  of  a  round  retort  having 
tuyeres  at  the  top  where  all  of  the  air  for  combustion  is  admitted.  Coal 
is  fed  from  a  storage  hopper  outside  of  the  boiler  by  means  of  a  worm  into 
the  bottom  of  this  retort  and  beneath  the  fire.  The  equipment  includes  a 
blower  which  is,  driven  by  the  same  motor  that  drives  the  stoker. 

Some  domestic  stokers  are  provided  with  automatic  grate  shaking 
.mechanism  together  with  screw  conveyers  for  removing  the  ash  frorn>  the 

458 


CHAPTER  28 — AUTOMATIC  FUEL  BURNING  EQUIPMENT 

ashpit  and  depositing  it  in  an  ash  receptacle  outside  the  boiler.  Certain 
types  can  also  be  provided  with  a  coal  conveyer  which  takes  coal  from  the 
storage  bin  and  maintains  a  full  hopper  at  the  stoker.  They  may  feed 
coal  to  the  furnace  either  intermittently  or  with  a  continuous  flow  regu- 
lated automatically  to  suit  conditions.  Where  the  boiler  is  provided  with 
indirect  coils  for  heating  the  domestic  hot  water,  the  stoker  may  be  so 
arranged  that  it  can  be  used  the  entire  year  to  maintain  a  continuous  hot 
water  supply. 

Class  2  Stokers,  Apartment  House 

This  class  is  used  extensively  for  heating  plants  in  apartments  and 
hotels,  and  also  for  small  industrial  plants  such  as  laundries,  bakeries,  and 
creameries.  The  various  stokers  in  this  class  differ  materially  in  their 
design,  although  the  majority  are  of  the  underfeed  type.  The  principal 
exception  is  an  overfeed  type  having  step  action  grates  in  a  horizontal 
plane  and  so  arranged  that  they  are  alternately  moving  and  stationary, 
and  are  designed  to  advance  the  fuel  during  combustion  to  an  ash  plate 
at  the  rear. 

All  of  the  stokers  are  provided  with  a  coal  hopper  outside  of  the  boiler. 
In  the  underfeed  types,  the  coal  feed  from  this  hopper  to  the  furnace  may 
be  accomplished  by  a  continuously  revolving  worm  or  by  an  intermittent 
plunger.  The  drive  for  the  coal  feed  may  be  an  electric  motor,  or  a  steam 
or  hydraulic  cylinder.  With  an  electric  motor,  the  connection  between 
the  driver  and  the  coal  feed  may  be  through  a  variable  speed  gear  train 
which  provides  two  or  more  speeds  for  the  coal  feed ;  or  it  may  be  through 
a  simple  gear  train  and  a  variable  speed  driver  for  the  change  in  speed  of 
the  coal  feed ;  or  a  simple  gear  train  with  a  coal  feed  having  an  adjustment 
for  varying  the  travel  of  the  feeding  device.  With  a  steam  or  hydraulic 
cylinder,  the  power  piston  is  connected  directly  to  the  coal  feeding 
plunger. 

.  The  stokers  in  this  class  vary  also  in  their  retort  design.  It  is  customary 
in  the  worm-feed  type  to  use  a  short  retort  in  order  that  the  unsupported 
length  of  worm  within  the  retort  may  not  be  too  weak  for  continuous 
service.  In  this  type  the  retort  is  placed  approximately  in  the  middle  of 
the  furnace  and  is  provided  with  tuyere  openings  at  the  top  on  all  sides. 
In  the  plunger-feed  type  the  retort  extends  from  the  inside  of  the  front 
wall  entirely  to  the  rear  wall  or  to  within  a  short  distance  of  the  rear  wall. 
This  type  of  retort  has  tuyeres  on  the  sides  and  at  the  rear. 

This  class  of  stokers  also  differs  in  the  grate  surface  surrounding  the 
retort.  In  many  of  the  worm-feed  stokers  this  grate  is  entirely  a  dead 
plate  on  which  the  fuel  rests  while  combustion  is  completed.  In  the  dead- 
plate  type,  all  of  the  air  for  combustion  is  furnished  by  the  tuyeres  at  the 
retort.  Because  of  this,  combustion  is  well  advanced  over  the  retort  so 
that  it  may  easily  be  completed  by  the  air  which  percolates  through  the 
fuel  bed.  With  the  dead-plate  type  of  grate  the  ash  is  removed  through 
the  fire  doors  and  it  is  therefore  desirable  that  the  fuel  used  shall  be  one 
in  which  the  ash  is  readily  reduced  to  a  clinker  at  the  furnace  temperature, 
in  order  that  it  may  be  removed  with  the  least  disturbance  of  the  fuel  bed. 

,  In  other  stokers  in  this  class,  the  grates  outside  ,of  the  retort  are  air- 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

admitting  and  shaking  grates.  These  grates  permit  a  large  part  of  the  ash 
to  be  shaken  into  the  ash  pit  beneath,  while  the  clinkers  are  removed 
through  the  fire  doors.  With  this  type  of  grate,  the  main  air  chamber 
extends  only  under  the  retort  while  the  side  grates  receive  air  by  natural 
draft  from  the  ash  pit. 

In  still  other  stokers  of  this  class,  the  main  air  chamber  extends  beyond 
the  retort  and  is  covered  with  fuel-bearing,  air-supplying  grates.  With 
this  type  of  grate,  the  fuel  is  supplied  with  air  from  the  main  air  chamber 
throughout  combustion.  Also  with  this  type  of  grate,  dump  plates  are 
provided  beyond  the  grates  where  the  ash  accumulates  and  from  which  it 
can  be  dropped  periodically  into  the  ash  pit  beneath. 

Stokers  in  this  class  are  compactly  built  in  order  that  they  may  fit  into 
standard  heating  boilers  and  still  leave  room  for  sufficient  combustion 
space  above  the  grates.  The  height  of  the  grate  is  approximately  the  same 
as  that  of  the  ordinary  grates  of  boilers,  so  that  it  is  usually  possible  to 
install  such  stokers  with  but  minor  changes  in  the  existing  equipment. 
In  some  districts,  there  are  statutory  regulations  governing  such  settings. 

These  stokers  vary  in  furnace  dimensions  from  30  in.  square  to  approxi- 
mately 66  in.  square.  The  capacity  of  the  stokers  is  measured  by  the 
amount  of  coal  that  can  be  burned  per  hour.  In  general,  manufacturers 
recommend  that,  for  continuous  operation,  the  coal  burning  rate  shall  not 
exceed  25  Ib  of  coal  per  square  foot  of  grate  per  hour,  while  for  short 
peaks  this  rate  may  be  increased  to  30  Ib  per  hour.  Although  these 
stokers  were  designed  to  burn  bituminous  coal,  they  can  also  be  used  to 
burn  the  small  sizes  of  anthracite  but  at  a  somewhat  lower  rate.  It  is 
often  customary  to  have  the  janitor  or  some  other  attendant  care  for  the 
boiler  as  one  of  his  duties.  Under  these  conditions  the  heating  plant  does 
not  receive  the  same  careful  attention  as  it  would  if  a  man  devoted  his 
entire  attention  to  the  fire.  With  periodic  hand-firing,  the  boiler  is 
operated  inefficiently  much  of  the  time.  With  a  stoker,  the  boiler  is 
operated  at  the  rate  that  the  conditions  require  so  long  as  there  is  coal  in 
the  hopper.  With  hand  firing,  it  is  customary  to  use  the  more  expensive 
sizes  of  fuel,  while  with  a  stoker  the  smaller  sizes  are  used  at  a  considerable 
saving  in  the  cost  per  ton.  Because  the  stoker  responds  promptly  to  auto- 
matic regulation,  it  is  possible  to  maintain  a  reasonably  constant  standard. 
Also  because  the  stoker  feeds  the  fuel  regularly  and  in  small  quantities 
without  losses  due  to  opening  doors,  it  must  of  necessity  be  more  efficient 
than  hand  firing.  This  increase  in  efficiency  depends  entirely  on  cbn- 
ditions,  with  a  minimum  of  about  10  per  cent  and  a  maximum  of  about 
25  per  cent. 

Class  3  Stokers,  General  Commercial 

These  stokers  are  suitable  for  the  heating  plants  of  large  schools,  hotels, 
hospitals,  or  other  large  institutions  as  well  as  industrial  plants.  This 
class  is  served  both  by  overfeed  stokers  and  by  underfeed  stokers.  The 
overfeed  stokers  are  in  general  of  three  types,  (1)  the  chain  grate,  (2) 
the  rear  cleaning  inclined  grate,  and  (3)  the  center  cleaning  inclined  grate 
or  V-type. 

Stokers  of  this  type  are  usually  operated  by  natural  draft,  although  in 
some  cases  conditions  permit  the  operation  of  forced  draft  under  the 

460 


CHAPTER  28 — AUTOMATIC  FUEL  BURNING  EQUIPMENT 

grates.  With  most  fuels,  it  is  not  advisable  to  operate  overfeed  grates  at 
too  high  a  combustion  rate  because  of  the  greater  difficulty  of  cleaning  and 
the  higher  maintenance,  but  where  the  fuel  is  free  burning  and  has  a  high 
ash'  fusion  temperature,  the  combustion  rate  is  not  so  restricted.  The 
operation  of  the  chain  grates  and  the  rear  cleaning  type  of  inclined  grates 
has  already  been  described. 

The  V-type  stoker  is  practically  obsolete  although  many  are  still  in 
operation.  In  this  stoker,  the  grates  are  inclined  downward  from  both 
sides  of  the  furnace  to  a  low  point  at  the  middle  where  there  is  either  a 
dump  plate  for  periodic  disposal  of  the  ash  or  a  rotary  ash  grate  for  con- 
tinuous discharge  of  ash.  In  this  stoker,  the  fuel  is  fed  into  a  hopper  at  the 
top  of  the  grate  on  each  side  of  the  furnace  and  advanced  down  the  grates 
to  the  center  where  the  refuse  is  accumulated.  This  stoker  is  always 
provided  with  a  combustion  arch  over  the  entire  furnace  for  the  purpose 
of  assuring  thorough  combustion  of  the  solid  fuel  and  providing  a  furnace 
temperature  sufficiently  high  to  burn  the  volatile  gases.  Because  of  this 
high  furnace  temperature  and  because  so  little  of  the  boiler  surface  is 
exposed  to  the  fire  to  assist  in  carrying  off  the  heat  by  radiation,  this 
stoker  is  characterized  by  severe  clinkering  in  the  ash  area.  With  all 
types  of  overfeed  stokers,  the  most  desirable  installations  are  in  boilers 
which  are  operated  with  comparatively  uniform  loads  and  moderate  rates 
of  combustion,  since,  even  with  good  combustion  arches,  fluctuating  loads 
or  high  combustion  rates  result  in  free  volatile  gas  and  this  in  turn  means 
smoke. 

The  underfeed  stokers  in  this  class  were  the  first  of  the  type  to  be 
developed  as  at  the  time  of  their  development  very  few  large  boilers  were 
in  use.  The  stokers  are  not  so  varied  in  design  as  those  in  the  smaller 
class  although  in  principle  they  are  much  the  same.  Practically  all  of 
them  are  of  the  plunger  coal  feed  type  with  retorts  extending  the  entire 
length  of  the  furnace,  with  air  supplying  grates  adjacent  to  the  retorts, 
and  with  manually-operated  dump  plates  at  the  sides  of  the  furnace. 
The  coal  feeding  plunger  is  operated  by  a  steam  or  electric  driver  through 
a  reduction  gearing,  or  by  a  steam  or  hydraulic  piston  connected  directly 
to  the  coal  feeding  plimger. 

These  stokers  are  heavily  built  and  designed  to  operate  continuously  at 
'high  boiler  ratings  with  a  minimum  amount  of  attention.  Because  of  the 
fact  that  all  volatile  gas  must  pass  through  the  fire  before  reaching  the 
combustion  chamber,  these  stokers  will  operate  smokelessly  under 
ordinary  conditions.  Also  because  of  the  fact  that  these  stokers  are 
always  provided  with  forced  draft,  they  are  the  most  desirable  type  for 
fluctuating  loads  or  high  boiler  ratings. 

In  the  design  of  the  grates  for  supporting  the  fuel  between  the  retort 
and  the  ash  plates,  the  stokers  differ  in  providing  for  movement  of  the 
fuel  during, combustion.  Some  stokers  are  designed  with  fixed  grates  of 
sufficient  angle  to  provide  for  this  movement  as  the  bed  is  agitated  by  the 
incoming  fuel,  while  others  have  alternate  moving  and  stationary  bars  in 
this  area  and  provide  for  this  movement  mechanically.  In  either  type, 
with  proper  operation,  all  refuse  will  be  deposited  at  the  dump  plate. 
Another  difference  in  these  stokers  is  that  some  makes  use  a  single  air 
chamber  under  the  whole  grate  area  thus  having  the  same  air  pressure 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

tinder  the  ignition  area  as  under  the  rest  of  the  grate,  while  others  have  a 
divided  air  chamber  using  the  full  air  pressure  under  the  ignition  area  and 
a  reduced  air  pressure  under  the  remainder  of  the  grate.  These  stokers 
vary  in  size  from  approximately  5  ft  square  to  a  maximum  of  8 J^  ft  square. 

Class  4  Stokers,  Large  Commercial 

These  stokers  are  usually  of  the  underfeed  type  with  multiple  retorts 
and  either  side  cleaning  or  rear  cleaning.  In  the  side  cleaning  type  there 
may  be  as  many  as  three  retorts  in  the  furnace,  and  the  stoker  functions 
in  the  same  manner  as  has  been  described  for  the  single  retort.  These 
stokers  ^are  usually  limited  in  length  to  approximately  8J^  ft  while  the 
width  may  be  as  great  as  10}^  ft.  In  the  rear  cleaning  stokers  the  number 
of  retorts  and  the  dimensions  of  the  furnace  are  practically  unlimited. 

DOMESTIC  OIL  BURNERS 

The  number  of  combinations  of  the  characteristic  elements  of  domestic 
oil  burners  is  rather  large  and  accounts  for  the  variety  of  burners  found  in 
actual  practice.  Domestic  oil  burners  may  be  classified  as  follows: 

1.  AIR  SUPPLY  FOR  COMBUSTION 

a.  Atmospheric — by  natural  chimney  draft. 

b.  Mechanical — electric-motor-driven  fan  or  blower. 

c.  Combination  of  (a}  and  (b) — primary  air  supply  by  fan  or  blower  and  secondary 

air  supply  by  natural  chimney  draft. 

2.  METHOD  OF  OIL  PREPARATION 

a.  Vaporizing — oil  distills  on  hot  surface  or  m  hot  cracking  chamber. 

b.  Atomizing — oil  broken  up  into  minute  globules. 

(1)  Centrifugal — by  means  of  rotating  cup  or  disc. 

(2)  Pressure — by  means  of  forcing  oil  under  pressure  through  a  small 

nozzle  or  orifice. 

(3)  Air  or  steam — by  high  velocity  air  or  steam  jet  in  a  special  type  of 

nozzle. 

(4)  Combination  air  and  pressure — by  air  entrained  with  oil  under  pressure 

and  forced  through  a  nozzle. 

c.  Combination  of  (a)  and  (b). 

3.  TYPE  OF  FLAME 

a.  Luminous — a  relatively  bright  flame.    An  orange-colored  flame  is  usually  best 

if  no  smoke  is  present. 

b.  Non-luminous — Bunsen-type  flame  (i.e.,  blue  flame). 

4.  METHODS  OF  IGNITION 

a.  Electric, 

(1)  Spark — by    transformer    producing    high-voltage    sparks.      Usually 

shielded  to  avoid  radio  interference.    May  take  place  continuously 
while  the  burner  is  operating  or  just  at  the  beginning  of  operation. 

(2)  Resistance — by  means  of  hot  wires  or  plates. 

b.  Gas. 

(1)  Continuous — pilot  light  of  constant  size. 

(2)  Expanding — size  of  pilot  light  expanded  temporarily  at  the  beginning 

of  burner  operation. 

462 


CHAPTER  28 — AUTOMATIC  FUEL  BURNING  EQUIPMENT 

c.  Combination — electric  sparks  light  the  gas  and  the  gas  flame  ignites  the  oil. 

d.  Manual — by  manually-operated  gas  torch  for  continuously  operating  burners. 

5.  MANNER  OF  OPERATION 

a.  On  and  off — burner  operates  only  a  portion  of  the  time  (intermittent). 

b.  High  and  low — burner  operates  continuously  but  varies  from  a  high  to  a  low 

flame. 

c.  Graduated — burner  operates  continuously  but  flame  is  graduated  according  to 

needs  by  regulating  both  air  and  oil  supply. 

Air  and  Oil  Supply 

The  object  of  regulating  the  air  and  oil  supplies  is  to  obtain  a  complete 
mixture  of  the  proper  quantities  of  oil  and  air  so  the  fire  will  be  clean  and 
efficient.  Proper  and  dependable  ignition  also  depends  upon  the  ability 
of  the  burner  to  produce  consistent  fuel-air  mixtures.  The  type  and  shape 
of  flame  depend  largely  upon  the  methods  of  air  and  oil  supply  employed. 

It  should  be  pointed  out  that  this  mixture  burns  in  a  space  called  the 
furnace,  which  is  lined  with  refractory  bricks  or  other  heat-resistant 
substances  for  the  purpose  of  maintaining  that  space  at  a  high  tempera- 
ture so  that  the  oil  and  air  may  completely  unite  and  burn.  Excessive 
cooling  before  combustion  is  completed  stops  the  combustion  process  and 
causes  soot.  The  furnace  is  in  some  instances  a  valuable  auxiliary  in 
assisting  in  the  actual  mixture  of  oil  and  air  and  in  modifying  the  flame 
shape,  besides  its  primary  function  of  maintaining  high  temperatures. 
The  size  and  shape  of  furnace  required  are  important,  especially  where  the 
dimensions  of  the  space  into  which  the  burner  is  to  be  placed  are  already 
fixed. 

Atornization 

The  purpose  of  atomization  is  greatly  to  increase  the  surface  area  of  a 
given  quantity  of  oil  in  order  to  accelerate  the  change  from  the  liquid 
state  (in  which  oil  cannot  burn)  to  the  gaseous  or  vaporous  state,  in  which 
state  it  is  one  of  the  elementary  fuels,  gaseous  hydrocarbon.  This  con- 
version is  largely  accomplished  through  the  action  of  radiant  heat  energy 
upon  the  flying  globules  of  oil,  and  the  tremendously  increased  surface 
provided  aids  gasification. 

Air  for  Combustion 

Air  for  combustion  usually  is  supplied  by  a  motor-driven  fan,  several 
types  being  in  common  use.  Electric  motors  varying  from  Ko  hp  to  H  hp 
ar,e  used  and  are  started  and  stopped  by  the  control  mechanism.  In  most 
cases,  they  are  direct-coupled  to  the  fan  as  well  as  to  a  gear  or  lobe  pump 
for  drawing  the  oil  from  the  storage  tank,  and  in  some  cases,  to  a  pump 
for  forcing  the  oil  through  the  nozzle. 

All  of  the  air  required  for  combustion  can  be  supplied  by  the  blower,  or 
else  only  the  primary  air  can  be  supplied  under  pressure  and  provision 
rnadle  so  that  the  remainder  will  be  drawn  into  the  combustion  chamber 
by  the  natural  draft  developed  by  the  chimney  or  by  an  injector-like 
action  of  the  primary  air.  In  any  event  there  should  be  definite  control 
Of  the'  quantity  of  "air  as  well  as  of  the  rate  of  oil  supply.  Some  method  of 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

draft  regulation  is  advisable  in  order  to  secure  proper  air  regulation.  It  is 
necessary  to  supply  more  air  than  is  actually  required  for  complete  com- 
bustion of  the  oil,  but  the  amount  of  excess  air  should  be  reduced  to  the 
lowest  workable  minimum.  Laboratory  tests  frequently  show  25  per  cent 
to  50  per  cent  more  air  than  is  required  for  combustion,  yet  field  tests 
indicate  that  the  average  burner  operates  with  from  50  per  cent  to  125 
per  cent  excess  air,  with  a  corresponding  reduction  in  the  efficiency  of  the 
burner.  Many  domestic  burners  are  extremes  of  simplicity,  the  only  mov- 
ing parts  being  the  motor  armature  with  a  shaft  and  direct-connected 
fan  and  pump  set. 

Type  of  Flame,  Ignition 

If  the  vaporizing  of  the  atomized  oil  and  the  combustion  are  concurrent 
events,  a  luminous  flame  will  usually  result.  If  vaporization  and  mixing 
are  accomplished  before  combustion,  a  non-luminous  flame  will  result. 
Some  burners  may  produce  either  type  of  flame  according  to  the  adjust- 
ment made. 

A  limited  comparison  of  these  two  types  of  flame  shows  no  inherent 
superiority  of  one  over  the  other  so  far  as  thermal  efficiencies  are  concerned. 
This  is  definitely  true  when  the  burners  are  placed  in  boilers  having  ample 
indirect  surface,  and  is  probably  true  in  general.  The  moot  question  pf 
radiation  has  not  been  conclusively  settled.  There  are  indications  that 
the  radiation  of  luminous  and  non-luminous  flames  in  boiler  furnaces  are 
practically  the  same.  More  information  is  needed  upon  this  subject. 

It  is  true,  however,  that  a  non-luminous  flame  may  show  low  excess  air 
and  the  presence  of  carbon  monoxide,  but  no  smoke.  Low  excess  air 
with  a  luminous  flame  will  usually  show  little  or  no  carbon  monoxide,  but 
will  be  unmistakably  smoky.  Visual  indications,  especially  with  a  blue 
flame,  may  therefore  be  quite  unreliable. 

When  a  burner  is  operating  intermittently  under  the  control  of  a 
thermostat,  some  positive  form  of  ignition  is  required  to  function  every 
time  there  is  a  call  for  heat. 

The  necessity  for  certain  ignition  under  adverse  conditions,  when  the 
line  voltage  is  low  or  the  oil  is  cold  is  paramount,  and  this  phase  of  burner 
design  and  operation  has  been  given  the  closest  attention,  as  faulty  igni- 
tion is  more  to  be  feared  than  improper  operation  once  the  flame  is 
established. 

The  effect  of  the  air  and  oil  setting  is  important,  since  it  may  be  neces- 
sary in  some  instances  to  adjust  for  greater  excess  air  than  is  otherwise 
required  in  order  to  get  a  mixture  suitable  for  certainty  of  ignition. 

Recent  research1  at  Yale  University,  conducted  in  cooperation  with  the 
A.S.H.V.E.  Research  Laboratory  and  the  American  Oil  Burner  Associa- 
tion, reveals  that  the  various  methods  of  operation  (i.e.,  on  and  off,  high 
and  low,  and  graduated)  all  have  potential  advantages  and  disadvantages 
and  that  a  choice  in  any  case  requires  a  consideration  of  the  heat-absorbing 
characteristics  of  the  boiler  in  which  the  burner  is  to  operate.  From  the 
standpoint  of  efficiency  of  operation  it  seems  that  there  is  little  choice  if 


^Intermittent  Operation  of  Oil  Burners,  by  L.  E.  Seeley  and  J.  H.  Powers  (A.S.H.V.E.  TRANSACTIONS, 
Vol.  38,  1932). 

464 


CHAPTER  28 — AUTOMATIC  FUEL  BURNING  EQUIPMENT 

at  the  maximum  setting  of  all  burners  the  boiler  efficiency  were  at  its 
maximum  value.  If,  on  the  other  hand,  the  boiler  were  operating  beyond 
its  point  of  maximum  efficiency,  then  it  appears  that  the  graduated  type 
might  show  better  results.  The  following  additional  factors  should  be 
considered : 

1.  The  intermittent  type  should  be  set  at  its  point  of  most  efficient  combustion. 

2.  An  intermittent  burner  should  be  set  for  a  higher  total  heat  output  than  either  of 
the  other  two  types  in  order  to  get  the  acceleration  necessary  when  heat  is  required. 
This  may  in  some  cases  give  less  economical  performance  due  to  the  increased  boiler  load. 

3.  An  interruption  in  electric  current  might  in  some  instances  be  troublesome  with 
continuously  operating  burners  where  manual  ignition  is  employed. 

4.  The  continuously  operating  burners  must  have  a  minimum  fuel  setting  low  enough 
to  prevent  overheating  in  mild  weather  or  during  the  summer  if  the  boiler  is  used  for 
domestic  hot  water. 

5.  Electrical  operating  costs  must  be  considered,  but  must  be  based  upon  known 
power  requirements.    The  power  requirements  of  some  burners  will  be  several  times  as 
nigh  as  others,  so  any  generalization  on  operating  costs  is  futile. 

6.  Evenness  of  heat  supply  will  have  some  influence  on  uniformity  of  temperature. 

7.  Number  and  cost  of  controls  which  reflect  in  the  manufacturing  costs. 

This  entire  subject,  therefore,  is  likely  to  be  somewhat  perplexing 
because  of  the  necessity  of  knowing,  and  the  difficulty  in  determining, 
the  efficiency  characteristics  of  many  heating  boilers.  Selections  of  oil 
burners  on  the  basis  of  their  manner  of  operation  will  probably  be  largely 
a  matter  of  preference.  The  advent  of  special  boilers  for  oil  burning  will 
provide  the  engineer  with  the  opportunity  for  greater  discrimination. 

Temperature  Control,  Protective  Devices 

Domestic  oil  burners  are  controlled  directly  from  the  change  in  tem- 
perature of  a  designated  control  room  (usually  the  living  room,  dining 
room  or  hall),  and  by  temperature  or  pressure  variations  in  the  boiler. 
Oil-burner  installations  put  in  only  a  few  years  ago  were  simplified  to 
the  extent  of  having  a  single  control  element — the  room  thermostat — 
that  started  and  stopped  the  burner.  The  modern  installation  provides, 
in  addition,  electrical  devices  inter- wired  with  the  control  system  to  in- 
sure against  poor  operation  and  to  guard  against  troubles  brought  on 
by  the  characteristics  of  the  heating  plant. 

One  control  system  provides  an  instrument  actuated  by  two  tempera- 
ture bulbs,  one  placed  in  the  outdoor  air  and  the  other  in  a  designated 
part  of  the  heating  system.  The  control  is  actuated  by  both  bulbs  and  is 
designed  to  maintain  the  heating  medium  at  a  temperature  to  suit  the 
variations  in  outdoor  temperature,  the  lower  the  outdoor  temperature  the 
higher  the  temperature  of  the  heating  medium.  Other  devices  have  been 
developed  to  maintain  a  certain  minimum  temperature  that  will  effectively 
prevent  the  downward  window  currents  of  cold  air  from  reaching  and 
traveling  across  the  floor,  regardless  of  the  room  thermostat. 

Owing  to  the  comparative  intensity  of  heat  production  with  a  burner, 
a  boiler  with  limited  water  storage  above  the  crown  sheet  might  pass 
steam  to  the  radiator  system  so  rapidly,  at  starting,  that  the  sheet  would 
be  uncovered,  with  probable  damage  to  the  boiler  structure.  A  low-water 
safety  can  be  so  wired  into  the  system  that  the  burner  will  be  stopped 
before  the  water  level  is  reduced  to  the  danger  point,  or  a  boiler  feed  can 

465 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

be  installed  to  add  water  to  the  boiler  to  maintain  a  safe  level,  instead 
of  stopping  the  burner.  Either  or  both  should  form  part  of  a  first-class 
installation. 

Again,  with  either  steam  or  water  systems,  the  burner  control  can  be 
inter-wired  with  a  thermostatic  device  having  its  temperature  Clement 
introduced  into  the  boiler  near  the  top,  its  function  being  to  limit  the 
maximum  temperature  of  water  or  pressure  of  steam  so  the  burner  will 
be  shut  off  before  dangerous  temperatures  or  pressures  are  reached.  Win- 
dows of  the  room  in  which  the  thermostat  is  located  are  sometimes  opened 
to  air  out  the  house  in  the  morning,  and  if  they  are  not  closed  promptly, 
the  burner  will  operate  continuously  and  possibly  develop  temperature 
and  pressure  conditions  that  might  be  detrimental  to  the  boiler.  This  is 
where  the  safety  device  can  be  used  to  offset  the  carelessness  of  the 
human  being. 

Safety  controls  have  been  developed  for  intermittent  burners  to  guard 
against  failure  of  ignition  and  in  some  instances  against  momentary  flame 
failures.  In  general,  regulatory  devices  are  well  developed  and  depend- 
able. Otherwise  the  domestic  oil  burner  probably  would  not  have  been 
possible. 

For  further  information  on  temperature  control  with  oil  burners,  see 
Chapter  14. 

Boilers  for  Domestic  Oil  Burners2 

Boilers  used  with  domestic  installations  may  be  those  designed  for  solid 
fuel  or  those  designed  for  liquid  fuel.  The  latter  are  coming  to  the  fore 
with  great  rapidity  as  they  usually  have  greatly  increased  secondary  sur- 
face. Many  are  of  copper  or  steel  tube  design.  Increased  efficiencies  of 
5  to  15  per  cent  are  often  obtainable  with  boilers  designed  especially  for 
liquid  fuel. 

It  is  possible  to  go  to  extremes  in  providing  secondary  surfaces  sufficient 
to  reduce  flue  temperatures  to  the  order  of  250  F  to  300  F,  with  the  result 
that  the  added  resistance  through  the  flues  may  necessitate  the  use  of  a 
booster  fan  to  insure  sufficient  draft.  It  is  difficult  to  obtain  satisfactory 
efficiencies  with  boilers  having  little  or  no  secondary  surfaces,  where  the 
hot  products  of  combustion  pass  almost  immediately  from  the  combustion 
chamber  to  the  flue;  in  fact  a  high  efficiency  is  unlikely  with  any  fuel 
under  such  conditions,  and  the  intermittent  burner  is  especially  at  a 
disadvantage  because  of  its  characteristic  development  of  heat  at  a  high 
rate  while  it  is  operating. 

It  is  essential  that  the  flame  produced  by  an  oil  burner,  especially  where 
it  is  strongly  luminous,  be  kept  from  contact  with  the  water-backed  sur- 
faces of  the  combustion  chamber,  and  to  this  end  bricking  or  its  equivalent 
must  be  provided  in  most  cases.  Where  a  burner  fires  through  the  ash  pit, 
doorframe  bricking  must  protect  the  unbacked  surfaces  of  the  ash  pit. 
The  same  fire  bricking  constitutes  the  actual  combustion  chamber  for  the 
burner  flame,  and  materially  increases  the  combustion  volume  for  a  given 
boiler. 


*For  additional  information  on  this  subject,  refer  to  Study  of  Performance  Characteristics  of  Oil  Burners 
and  Low-Pressure  Heating  Boilers,  by  L.  E.  Seeley  and  E,  J.  Tavanlar  (A.S.H.V.E.  TRANSACTIONS,  Vol.  37, 
1931). 

466 


CHAPTER  28 — AUTOMATIC  FUEL  BURNING  EQUIPMENT 

Installation 

The  intelligence  and  care  with  which  a  burner  is  installed  largely  deter- 
mine the  satisfaction  that  will  result  from  its  operation.  Two  plans  for  the 
installation  of  burners  are  in  general  use.  In  the  first,  the  dealer  makes 
all  installations.  In  the  other,  sales  agencies  function  only  to  make  sales, 
and  the  installation  for  as  many  as  twenty  such  sales  offices  is  done  by  a 
centrally  located  installation  force,  usually  factory  controlled. 

Some  burners  are  adjusted  for  oil  rate  by  means  of  a  blind  needle  valve 
that  can  be  operated  only  with  a  special  wrench ;  others,  by  changing  the 
size  of  the  orifice;  others,  by  a  combination  of  orifice  size  and  pressure. 
In  any  event,  changes  in  the  firing  rate,  involving  careful  air  and  draft 
adjustment  to  match  the  oil  rate,  should  be  made  by  only  a  trained  man, 


to. 


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FIG.  1.    FULL  LOAD  RATE  OF  OIL  CONSUMPTION  FOR  HEATING  BOILERS 


preferably  with  the  aid  of  an  Orsat  test  set  so  that  the  degree  of  combus- 
tion efficiency  can  be  determined.  It  is  practically  impossible  to  set  a 
burner  flame  by  eye,  although  that  has  been  general  practice  in  the  past. 
The  industry  is  turning  to  the  Orsat  and,  as  a  result,  more  domestic 
burners  are  operating  at  from  9  to  12  per  cent  C02,  representing  a  higher 
efficiency  combustion  than  at  5  to  8  per  cent,  as  frequently  is  the  case 
where  the  burner  is  adjusted  by  eye. 

Air  for  Combustion 

It  is  essential  that  the  basement,  or  at  least  that  portion^  used  ^as  a 
boiler  room,  be  open  to  the  outside  air,  in  order  that  sufficient  air  be 
available  for  combustion.  Frequently  a  case  of  poor  operation  will  be 
found  where  a  test  with  a  draft  gage  made  by  inserting  the  tube  through 
the  keyhole  of  the  outer  door  will  show  that  there  is  a  partial  vacuum  in 
the  basement  when  the  burner  is  running,  all  of  the  combustion  air  coming 

467 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

through  the  keyhole  and  minute  cracks.  A  simple  remedy  is  to  cut  an 
inch  from  the  bottom  of  the  outer  door. 

In  order  to  achieve  satisfactory  heating  at  the  lowest  cost,  careful  con- 
sideration should  be  given  to  oil,  air  and  draft  adjustment.  The  oil 
adjustment  should  be  determined  from  the  total  heat  requirements  to  be 
met.  The  heat  loss  of  the  building  plus  an  allowance  for  piping  plus 
20  to  25  per  cent  for  pick-up  establishes  the  maximum  output  required 
from  the  boiler.  Fig.  1  indicates  the  oil  required  in  gallons.  ^  Piping 
allowances  will  usually  vary  between  25  and  10  per  cent,  decreasing  with 
an  increase  in  the  size  of  the  building. 

With  the  oil  rate  thus  fixed,  the  air  and  draft  should  be  set  to  give 
efficient  combustion  (that  is,  10  to  12  per  cent  C02).  The  furnace  draft 
should  be  set  reasonably  low  and  should  be  maintained  constant  by 
means  of  an  automatic  draft  regulator.  Without  this  the  air  supply  will 
fluctuate,  causing  uneven  performance.  A  check  should  be  made  to 
insure  that  ignition  will  be  satisfactory  under  all  conditions.  An  oil  burner 
of  the  continuous  type  might  dispense  with  all  or  part  of  the  pick-up 
allowance  due  to  the  nature  of  its  operation.  Careful  adjustment  will 
provide  ample  heat  output  under  all  conditions,  will  minimize  the  load  on 
the  boiler,  and  will  establish  the  most  favorable  conditions  for  intermit- 
tent operation. 

An  essential  element  in  the  satisfactory  operation  of  domestic  oil  burners 
is  the  provision  for  maintenance  and  service  for  the  burners.  What  might 
be  called  emergency  service  for  mechanical  or  electrical  failure  of  the  burner 
has  rapidly  diminished  during  the  last  few  years  until  a  level  has  been 
reached  where  groups  of  100  to  1000  burners  in  a  community  consistently 
will  require  an  average  of  not  more  than  one  call  per  burner  per  heating 
season.  Maintenance  service  is  coming  into  general  practice  where,  for  a 
fixed  annual  payment,  regular  inspection  is  made  of  the  burner,  and  faulty 
operation  corrected  before  the  burner  becomes  inoperative.  This  service 
may  contemplate  entire  overhauling  of  the  burner  during  the  summer, 
and  may  include  annual  cleaning  of  the  boiler  flues  with  a  specially  de- 
signed vacuum  cleaner. 

Domestic  Hot  Water  Supply 

Provision  may  be  made  for  heating  domestic  water  through  exchange 
heaters  attached  to  the  boiler,  in  which  water  is  maintained  at  a  fixed 
temperature  or  steam  at  a  set  pressure  during  the  entire  year.  The  flow 
of  water  or  steam  to  the  radiators  is  controlled  by  electrically-operated 
valves,  which  remain  closed  during  warm  weather  and  open  (through  the 
functioning  of  the  room  thermostat)  when  heat  is  required  in  the  house. 
The  room  thermostat  either  causes  heat  to  be  produced  by  starting  the 
burner  when  the  room  temperature  drops  to  a  predetermined  point,  or 
closes  the  circuit  of  the  motor  by  operating  a  valve  in  the  flow  line  of  the 
heating  system,  the  motor  opening  the  water  or  steam  valve  and  per- 
mitting water  or  steam  immediately  to  flow  to  the  radiators.  When  the 
flow  in  a  water  heating  system  is  sluggish,  the  room  thermostat  also  can 
start  the  motor  of  a  circulation  pump,  thereby  decreasing  the  time  re- 
quired to  bring  the  room  temperature  up  to  the  desired  point. 

It  is  usual  in  small  steam  heating  systems  to  dispense  with  the  motor- 

468 


CHAPTER  28 — AUTOMATIC  FUEL  BURNING  EQUIPMENT 

operated  valve  and  by  means  of  an  aquastat  maintain  the  boiler  water  at 
a  constant  temperature  but  well  below  the  steaming  temperature  (i.e., 
140  to  180  F).  The  lowest  temperature  setting  that  will  produce  suf- 
ficiently hot  water  will  be  the  most  economical.  The  aquastat  will 
always  function  in  such  a  way  as  to  maintain  this  temperature  except 
when  the  room  thermostat  calls  for  heat,  which  means  that  a  call  for 
steam  can  be  more  quickly  obtained. 

Another  type  of  control  valve  available  for  hot  water  systems  is 
thermostatically  operated  so  as  to  prevent  a  flow  of  water  to  the  heating 
system  until  the  call  of  the  room  thermostat  for  heat  raises  the  water 
temperature  above  that  normally  required  for  domestic  hot  water.  It 
should  be  noted  that,  except  in  the  case  of  the  graduated  burner,  the 
water  temperature  in  the  heating  system  will  nearly  always  reach  its 
maximum,  thereby  depriving  this  system  to  some  degree  of  its  natural 
advantage  of  modulation. 

COMMERCIAL  OIL  BURNERS 

Liquid  fuels  are  used  for  heating  apartment  buildings,  hotels,  public 
and  office  buildings,  schools,  churches,  hospitals,  department  stores,  as 
well  as  industrial  plants  of  all  kinds.  Contrary  to  domestic  heating,  con- 
venience seldom  is  a  dominating  factor,  the  actual  net  cost  of  heat  pro- 
duction usually  controlling  the  selection  of  fuel.  Some  of  the  largest  office 
buildings  have  been  using  oil  for  many  years.  Many  department  stores 
have  found  that  floor  space  in  basements  and  sub-basements  can  be  used 
to  better  advantage  for  merchandising  wares,  and  credit  the  heat  pro- 
ducing department  with  this  saving. 

Wherever  possible,  the  boiler  plant  should  be  so  arranged  that  either 
oil  or  solid  fuel  can  be  used  at  will,  permitting  the  management  to  take 
advantage  of  changes  in  fuel  costs  if  any  occur.  Each  case  should  be 
considered  solely  in  the  light  of  local  conditions  and  prices. 

Burners  for  commercial  heating  may  be  either  large  models  of  types 
used  in  domestic  heating,  or  special  types  developed  to  meet  the  condi- 
tions imposed  by  the  boilers  involved.  Generally  speaking,  such  burners 
are  of  the  mechanical  or  pressure  atomizing  types,  the  former  using 
rotating  cups  producing  a  horizontal  torch-like  flame.  As  much  as  350  gal 
of  oil  per  hour  can  be  burned  in  these  units,  and  frequently  they  are 
arranged  in  multiple  on  the  boiler  face,  from  two  to  five  burners  to  each 
boiler. 

The  larger  installations  are  nearly  always  started  with  a  hand  torch, 
and  are  manually  controlled,  but  the  use  of  automatic  control  is  increasing, 
and  completely  automatic  burners  are  now  available  to  burn  the  two 
heaviest  grades  of  oil.  Nearly  all  of  the  smaller  installations,  in  schools, 
churches,  apartment  houses  and  the  like,  are  fully  automatic. 

Because  of  the  viscosity  of  the  heavier  oils,  it  is  customary  to  heat  them 
before  transferring  by  truck  tank.  It  also  has  been  common  practice  to 
preheat  the  oil  between  the  storage  tank  and  the  burner,  as  an  aid  to 
movement  of  the  oil  as  well  as  to  atomization.  This  heating  is  accomplished 
by  heat-transfer  coils,  using  water  or  steam  from  the  heating  boiler,  and 
heating  the  oil  to  within  30  deg  of  its  flash  point. 

469 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

Unlike  the  domestic  burner,  units  for  large  commercial  applications 
frequently  consist  of  atomizing  nozzles  or  cups  mounted  on  the  boiler 
front  with  the  necessary  air  regulators,  the  pumps  for  handling  the  oil 
and  the  blowers  for  air  supply  being  mounted  in  sets  adjacent  to  the 
boilers.  In  such  cases,  one  pump  set  can  serve  several  burner  units,  and 
common  prudence  dictates  the  installation  of  spare  or  reserve  pump  sets. 
Pre-heaters  and  other  essential  auxiliary  equipment  also  should  be  in- 
stalled in  duplicate. 

Boiler  Settings 

As  the  volume  of  space  available  for  combustion  is  the  determining 
factor  in  oil  consumption,  it  is  general  practice  to  remove  grates  and 
extend  the  combustion  chamber  downward  to  include  or  even  exceed  the 
ash-pit  volume;  in  new  installations  the  boiler  should  be  raised  to  make 
added  volume  available.  Approximately  1  cu  ft  of  combustion  volume 
should  be  provided  for  every  developed  boiler  horsepower,  and  in  this 
volume  from  1.5  to  2  Ib  of  oil  can  properly  be  burned.  This  cor- 
responds to  a  maximum  liberation  of  about  38,000  Btu  per  cubic  foot  per 
hour.  There  are  indications  that  at  times  much  higher  fuel  rates  may  be 
satisfactory.  This  in  turn  suggests  that  the  value  of  38,000  Btu  per  cubic 
foot  per  hour  might  be  adjusted  according  to  good  engineering  judgment. 
For  best  results,  care  should  be  taken  to  keep  the  gas  velocity  below  40  ft 
per  second.  Where  checkerwork  of  brick  is  used  to  provide  secondary  air, 

food  practice  calls  for  about  1  sq  in.  of  opening  for  each  pound  of  oil 
red  per  hour.  Such  checkerwork  is  best  adapted  to  flat  flames,  or  to 
conical  flames  that  can  be  spread  over  the  floor  of  the  combustion  chamber. 
The  proper  bricking  of  a  large  or  even  medium  sized  boiler  for  oil  firing  is 
important  and  frequently  it  is  advisable  to  consult  an  authority  on  this 
subject.  The  essential  in  combustion  chamber  design  is  to  provide 
against  flame  impingement  upon  either  metallic  or  fire-brick  surfaces. 
Manufacturers  of  oil  burners  usually  have  available  detailed  plans  for 
adapting  their  burners  to  various  types  of  boilers,  and  such  information 
should  be  utilized. 

GAS-FIRED  APPLIANCES 

The  increased  use  of  gas  for  house  heating  purposes  has  resulted  in  the 
production  of  such  a  large  number  of  different  types  of  gas-heating 
systems  and  appliances  that  today  there  is  probably  a  greater  variety  of 
them  than  there  is  for  any  other  kind  of  fuel. 

Gas-fired  heating  systems  may  be  classified  as  follows: 

I.    Gas-Designed  Heating  Systems. 

A.  Central  Heating  Plants. 

1.  Steam,  hot  water,  and  vapor  boilers. 

2.  Warm  air  furnaces. 

B.  Unit  Heating  Systems. 

1.  Warm  air  floor  furnaces. 

2.  Industrial  unit  heaters. 

3.  Space  heaters. 

4.  Garage  heaters. 

470 


CHAPTER  28 — AUTOMATIC  FUEL  BURNING  EQUIPMENT 

II.  'Conversion  Heating  Systems. 
A.  Central  Heating  Plants. 

1.  Steam,  hot  water  and  vapor  boilers. 

2.  Warm  air  basement  furnaces. 

The  majority  of  these  systems  are  supplied  with  either  automatic  or 
manual  control.  Central  heating  plants,  for  example!  whether  gas 
designed  or  conversion  systems,  may  be  equipped  with  room  temperature 
control,  push  button  control,  or  manual  control. 

Although  no  exact  rules  can  be  prescribed  as  to  the  field  best  covered  by 
each  of  the  foregoing  systems,  each  installation  will  have  problems  point- 
ing more  or  less  directly  to  some  particular  type  of  heating  equipment. 

Gas-Fired  Boilers 

Information  on  gas-fired  boilers  will  be  found  in  Chapter  25. 

Either  snap  action  or  throttling  control  is  available  for  gas  boiler  opera- 
tion. This  is  especially  advantageous  in  straight  steam  systems  because 
steam  pressures  can  be  maintained  at  desired  points,  while  at  the  same 
time  complete  cut-off  of  gas  is  possible  when  the  thermostat  calls  for  it. 

Warm  Air  Furnaces 

There  are  two  general  classes  of  gas-fired  warm  air  furnaces,  the 
gravity  furnace  which  depends  upon  the  natural  tendency  of  heated  air  to 
rise,  providing  the  proper  circulation  of  heated  air  into  the  room,  and  the 
mechanical  circulation  furnace  by  which  the  air  to  be  heated  is  forced 
through  or  drawn  through  the  furnace  by  means  of  a  fan. 

Warm  air  furnaces  are  variously  constructed  of  cast  iron,  sheet  metal 
and  combinations  of  the  two  materials.  If  sheet  metal  is  used,  it  must  be 
of  such  a  character  that  it  will  have  the  maximum  resistance  to  the  cor- 
rosive effect  of  the  products  of  combustion.  With  some  varieties  of 
manufactured  gases,  this  effect  is  quite  pronounced.  Warm  air  furnaces 
are  obtainable  in  sizes  from  those  sufficient  to  heat  the  largest  residence 
down  to  sizes  applicable  to  a  single  room.  The  practice  of  installing  a 
number  of  separate  furnaces  to  heat  individual  rooms  is  peculiar  to  mild 
climates,  such  as  that  of  southern  California.  Small  furnaces,  frequently 
controlled  by  electrical  valves  actuated  by  push-buttons  in  the  room 
above,  are  often  installed  to  heat  rooms  where  heat  may  be  desired  for  an 
hour  or  so  each  day.  These  furnaces  are  used  also  for  heating  groups  of 
rooms  in  larger  residences.  In  a  system  of  this  type  each  furnace  should 
supply  a  group  of  rooms  in  which  the  heating  requirements  for  each  room 
in  the  group  are  similar  as  far  as  the  period  of  heating  and  temperature 
to  be  maintained  are  concerned.  Bedrooms,  living  rooms,  and  dining 
rooms  often  present  excellent  possibilities  for  this  type  of  furnace. 

The  same  fundamental  principle  of  design  that  is  followed  in  the  con- 
struction of  boilers,  that  is,  breaking  the  hot  gas  up  into  fine  ^  streams  so 
that  all  particles  are  brought  as  close  as  possible  to  the  heating  surface, 
is  equally  applicable  to  the  design  of  warm  air  furnaces.  The  desirability 
of  using  an  appliance  designed  for  gas,  when  gas  is  to  be  the  fuel,  applies 
even  more  strongly  to  furnaces  than  to  boilers. 

Codes  for  proportioning  warm  air  heating  plants,  such  as  that  formu- 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

lated  by  the  National  Warm  Air  Heating  Association  (see  note  p.  401), 
are  equally  applicable  to  gas  furnaces  and  coal  furnaces.  Recirculation 
should'  always  be  practiced  with  gas-fired  warm  air  furnaces.  It  not  only 
aids  in  heating,  but  is  essential  to  economy.  Where  fans  are  used  in 
connection  with  warm  air  furnaces  for  residence  heating,  it  is  well  to 
have  the  control  of  the  fan  and  of  the  gas  so  coordinated  that  there  will 
be  sufficient  delay  between  the  turning  on  of  the  gas  and  the  starting  of 
the  fan  to  prevent  blasts  of  cold  air  being  blown  into  the  heated  rooms. 
An  additional  thermostat  in  the  air  duct  easily  may  be  arranged  to 
accomplish  this. 

Floor  Furnaces 

Warm  air  floor  furnaces  are  well  adapted  for  heating  first  floors,  or 
where  heat  is  required  in  only  one  or  two  rooms.  A  number  may  be  used 
to  provide  heat  for  the  entire  building  where  all  rooms  are  on  the  ground 
floor,  thus  giving  the  heating  system  flexibility  as  any  number  of  rooms 
may  be  heated  without  heating  the  others.  With  the  usual  type  the 
register  is  installed  in  the  floor,  the  heating  element  and  gas  piping  being 
suspended  below.  Air  is  taken  downward  between  the  two  sheets  of  the 
double  casing  and  discharged  upward  over  the  heating  surfaces  and  into 
the  room.  The  appliance  is  controlled  from  the  room  to  be  heated  by 
means  of  a  control  lever  located  near  the  edge  of  the  register.  The  handle 
of  the  control  is  removable  as  a  precaution  against  accidental  turning 
on  or  off  of  the  gas  to  the  furnace. 

Space  heaters  are  generally  used  for  auxiliary  heating,  but  may  be,  and 
are  in  many  cases,  installed  for  furnishing  heat  to  entire  buildings.  Space 
heaters  are  quite  extensively  used  for  house  heating  in  milder  climates 
such  as  exist  in  the  South  and  Southwest.  With  the  exception  of  wall 
heaters,  they  are  portable,  and  can  be  easily  removed  and  stored  during 
the  summer  season.  Although  they  should  be  connected  with  solid  piping 
it  is  sometimes  desirable  to  connect  them  with  flexible  gas  tubing  in  which 
case  a  gas  shut-off  on  the  heater  is  not  permitted,  and  only  A.G.A. 
approved  tubing  should  be  used. 

Space  Heaters 

Parlor  furnaces  or  circulators  are  usually  constructed  to  resemble  a 
cabinet  radio.  They  heat  the  room  entirely  by  convection,  i.e.,  the  cold 
air  of  the  room  is  drawn  in  near  the  base  and  passes  up  inside  the  jacket 
around  a  drum  or  heating  section,  and  out  of  the  heater  at  or  near  the  top. 
These  heaters  cause  a  continuous  circulation  of  the  air  in  the  room  during 
the  time  they  are  in  operation.  The  burner  or  burners  are  located  in  the 
base  at  the  bottom  of  an  enclosed  combustion  chamber.  The  products  of 
combustion  pass  up  around  baffles  within  the  heating  element  or  drum, 
and  out  the  flue  at  the  back  near  the  top.  They  are  well  adapted  not  only 
for  residence  room  heating  but  also  for  stores  and  offices. 

Radiant  heaters  make  admirable  auxiliary  heating  appliances  to  be  used 
during  the  occasional  cool  days  at  the  beginning  and  end  of  the  heating 
season  when  heat  is  desired  in  some  particular  room  for  an  hour  or  two. 
The  radiant  heater  gives  off  a  considerable  portion  of  its  heat  in  the  form 
of  radiant  energy  emitted  by  an  incandescent  refractory  that  is  heated  by 

472 


CHAPTER  28 — AUTOMATIC  FUEL  BURNING  EQUIPMENT 

a  Bunsen  flame.  They  are  made  in  numerous  shapes  and  designs  and  in 
sizes  ranging  from  two  to  fourteen  or  more  radiants.  Some  have  sheet- 
iron  bodies  finished  in  enamel  or  brass  while  others  have  cast-iron  or  brass 
frames  with  heavy  fire  clay  bodies.  An  atmospheric  burner  is  supported 
near  the  center  of  the  base,  usually  by  set  screws  at  each  end.  Others 
have  a  group  of  small  atmospheric  burners  supported  on  a  manifold 
attached  to  the  base.  Most  radiant  heaters  are  supported  on  legs  and  are 
portable;  however,  there  are  also  types  which  are  encased  in  a  jacket 
which  fits  into  the  wall  with  a  grilled  front,  similar  to  the  ordinary  wall 
register.  Others  are  encased  in  frames  which  fit  into  fireplaces. 

Gas-fired  steam  and  hot  water  radiators  are  popular  types  of  room  heating 
appliances.  They  provide  a  form  of  heating  apparatus  for  intermittently 
heated  spaces  such  as  stores,  small  churches  and  some  types  of  offices  and 
apartments.  They  are  made  in  a  large  variety  of  shapes  and  sizes  and  are 
similar  in  appearance  to  the  ordinary  steam  or  hot  water  radiator  con- 
nected to  a  basement  boiler.  A  separate  combustion  chamber  is  provided 
in  the  base  of  each  radiator  and  is  usually  fitted  with  a  one-piece  burner. 
They  may  be  secured  in  either  the  vented  or  unvented  types,  and  with 
steam  pressure,  thermostatic  or  room  temperature  controls. 

Warm  air  radiators  are  similar  in  appearance  to  the  steam  or  hot  water 
radiators.  They  are  usually  constructed  of  pressed  steel  or  sheet  metal 
hollow  sections.  The  hot  products  of  combustion  circulate  through  the 
sections  and  are  discharged  out  a  flue  or  into  the  room,  depending  upon 
whether  the  radiator  is  of  the  vented  or  unvented  type. 

Garage  heaters  are  usually  similar  in  construction  to  the  cabinet 
circulator  space  heaters,  except  that  safety  screens  are  provided  over  all 
openings  into  the  combustion  chamber  to  prevent  any  possibility  of 
explosion  from  gasoline  fumes  or  other  gases  which  might  be  ignited  by 
an  open  flame.  They  are  usually  provided  with  automatic  room  tem- 
perature controls  and  are  well  suited  for  heating  either  residence  or 
commercial  garages. 

Conversion  Burners 

Residence  heating  with  gas  through  the  use  of  conversion  burners  in- 
stalled in  coal-designed  boilers  and  furnaces  represents  a  common  type 
of  gas-fired  house  heating  system,  especially  in  natural  gas  territories. 
In  many  conversion  burners  radiants  or  refractories  are  employed  to 
convert  some  of  the  energy  in  the  gas  to  radiant  heat.  Others  are  of  the 
blast  type  with  luminous  flames,  operating  without  refractories.  In  each 
case  an  attempt  is  made  to  transfer  the  majority  of  the  heat  from  the  gas 
to  the  medium  to  be  heated  within  the  fire  pot  itself  because  of  the  low 
heat  transfer  that  takes  place  in  the  flue  passages. 

Many  conversion  units  are  equipped  with  sheet  metal  secondary  air 
ducts  which  are  inserted  through  the  ash-pit  door.  The  duct  is  equipped 
with  automatic  air  controls  which  open  when  the  burners  are  operating 
and  close  when  the  gas  supply  is  turned  off.  This  prevents  a  large  part 
of  the  circulation  of  cold  air  through  the  combustion  space  of  the  ap- 
pliance when  not  in  operation.  By  means  of  this  duct  the  air  necessary 
for  proper  combustion  is  supplied  directly  to  the  burner,  thereby  making 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

it  possible  to  reduce  the  amount  of  excess  air  passing  through  the  com- 
bustion chamber. 

Conversion  units  are  made  in  many  sizes  both  round  and  rectangular 
to  fit  different  types  and  makes  of  boilers  and  furnaces.  They  may  be 
secured  with  manual,  push  button,  or  room  temperature  control. 

Sizing  Gas-Fired  Heating  Plants 

While  gas-burning  equipment  can  be  and  usually  is  so  installed  as  to  be 
completely  automatic,  maintaining  the  temperature  of  rooms  at  a  pre- 
determined and  set  figure,  there  are  in  use  installations  which  are  manually 
controlled.  Experience  has  shown  that  in  order  to  effectively  overcome 
the  starting  load  and  losses  in  piping,, a  manually-controlled  gas  boiler 
should  have  an  output  as  much  as  100  per  cent  greater  than  the  equivalent 
standard  cast-iron  column  radiation  which  it  is  expected  to  serve. 

Boilers  under  thermostatic  control,  however,  are  not  subject  to  such 
severe  pick-up  or  starting  loads.  Consequently,  it  is  possible  to  use  a 

TABLE  1.    SELECTION  FACTORS  FOR  GAS  BOILERS 


CAST-IRON  STEAM  RADIATION 
(EQUIVALENT  SQUARE  FEBT) 

SELECTION  FACTOR 
(PER  CENT) 

500 

56.0 

800 

54.0 

1,200 

51.0 

1,600 

48.0 

2,000 

45.0 

3,000 

42.5 

4,000  and  over 

40.0 

much  lower  selection,  or  safety,  factor.  A  gas-fired  boiler  under  ther- 
mostatic control  is  so  sensitive  to  variations  in  room  temperatures  that  in 
most  cases  a  factor  of  25  per  cent  is  sufficient  for  pick-up  load. 

The  factor  to  be  allowed  for  loss  of  heat  from  piping,  however,  must 
vary  somewhat,  the  proportionate  amount  of  piping  installed  being  con- 
siderably greater  for  small  installations  than  for  large  ones.  Consequently, 
a  selection  factor  for  thermostatically  controlled  boilers  must  be  variable. 
Table  1  gives  selection  factors  to  be  added  to  the  installed  steam  radiation 
under  thermostatic  control.  They  have  been  established  by  experience 
and  are  recommended  by  the  American  Gas  Association. 

The  same  factors  may  be  used  in  determining  the  gas  demand  for  which 
conversion  burners  installed  in  steam  or  hot  water  boilers  should  be  set. 
Multiplying  the  equivalent  direct  heating  surface  (radiation)  by  240  and 
adding  the  appropriate  percentage  from  Table  1,  and  then  dividing  by  the 
heat  value  of  the  gas  and  by  the  heating  efficiency  (see  discussion  of 
heating  efficiencies  in  Chapter  29),  gives  the  proper  hourly  rate  of  gas 
consumption.  However,  inadequate  boiler  heating  surface  for  gas 
burning,  often  encountered  in  coal-designed  boilers  converted  to  gas,  may 
necessitate  operation  at  a  lesser  demand,  resulting  in  much  slower  pick-up 
and  less  margin  of  safety  for  piping  loss. 

Appliances  used  for  heating  with  gas  should  bear  the  approval  seal1 

474 


CHAPTER  28 — AUTOMATIC  FUEL  BURNING  EQUIPMENT 

of  the  American  Gas  Association  Testing  Laboratory.  Installations  should 
be  made  in  accordance  with  the  recommendations  shown  in  the  publica- 
tions of  that  association. 

Ratings  for  Gas  Appliances 

Since  a  gas  appliance  has  a  heat-generating  capacity  that  can  be  pre- 
dicted accurately  to  within  1  or  2  per  cent,  and  since  this  capacity  is  not 
affected  by  such  things  as  condition  of  fuel  bed  and  soot  accumulation, 
makers  of  these  appliances  have  an  opportunity  to  rate  their  product  in 
exact  terms.  Consequently  all  makers  give  their  product  an  hourly  Btu 
output  rating.  This  is  the  amount  of  heat  that  is  available  at  the  outlet  of 
a  boiler  in  the  form  of  steam  or  hot  water,  or  at  the  bonnet  of  the  furnace 
in  the  form  of  warm  air.  The  output  rating  is  in  turn  based  upon  the 
Btu  input  rating  which  has  been  approved  by  the  American  Gas  Asso- 
ciation Testing  Laboratory  and  upon  an  average  efficiency  which  has 
been  assigned  by  that  association. 

In  the  case  of  boilers,  the  rating  can  be  put  in  terms  of  square  feet  of 
equivalent  direct  radiation  by  dividing  it  by  240  for  steam,  and  1503  for 
water.  This  gives  what  is  called  the  American  Gas  Association  rating,  and 
is  the  manner  in  which  all  appliances  approved  by  the  American  Gas 
Association  Laboratory  are  rated.  To  use  these  ratings  it  is  only  necessary 
to  increase  the  calculated  heat  loss  or  the  equivalent  direct  radiation  load 
by  an  appropriate  amount  for  starting  and  piping,  and  to  select  the  boiler 
or  furnace  with  the  proper  rating. 

The  rating  given  by  the  American  Gas  Association  Laboratory  is  not 
only  a  conservative  rating  when  considered  from  the  standpoint  of 
capacity  and  efficiency,  but  is  also  a  safe  rating  when  considered  from  the 
standpoint  of  physical  safety  to  the  owner  or  caretaker.  The  rating  that 
is  placed  upon  an  appliance  is  limited  by  the  amount  of  gas  that  can  be 
burned  without  the  production  of  harmful  amounts  of  carbon  monoxide. 
This  same  limitation  applies  to  all  classes  of  gas-consuming  heating 
appliances  that  are  tested  and  approved  by  the  Laboratory.  Gas  boilers 
are  available  with  ratings  up  to  14,000  sq  ft  of  steam,  while  furnaces  with 
ratings  up  to  about  500,000  Btu  per  hour  are  available.  (See  Chapter  23.) 

Installation  Features 

One  feature  of  the  piping  installation  that  adds  to  the  satisfactory 
service  rendered  by  gas  boilers  is  provision  for  adequate  and  rapid  venting 
of  the  air  from  steam  heating  systems.  If  air  leaks  into  the  steam  dis- 
tribution system  during  the  period  that  the  gas  is  turned  off,  and  then 
vents  out  slowly  when  the  thermostat  calls  for  heat,  the  result  will  be  a 
further  cooling  of  the  premises  between  the  time  that  the  thermostat 
calls  for  heat  and  the  time  that  steam  reaches  the  radiators.  A  freely 
venting  steam  or  vapor  system  gives  maximum  economy  and  minimum 
temperature  variation.  When  gas  boilers  are  attached  to  existing  heating 
plants,  it  is  good  practice  to  check  the  effectiveness  of  the  venting  devices 


»A  value  of  160  for  the  heat  emission  of  hot  water  radiators  is  used  by  many  engineers.  The  actual  heat 
emission,  however,  depends  on  the  temperature  of  the  water  and  of  the  surrounding  air.  See  Chapters 
30  and  33. 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

and  if  necessary  to  replace  them  with  more  effective  ones  that  will  prevent 
the  return  of  air  into  the  heating  system,  and  also  to  check  the  tightness 
of  the  piping. 

Frequently  when  a  coal  boiler  is  already  installed  in  a  home,  it 
is  expedient  to  leave  the  coal  boiler  in  place,  and  to  cross-connect  the 
gas  boiler  with  it.  Where  gas  heating  is  new  to  the  community,  it  pro- 
duces a  more  secure  feeling  in  the  customer's  mind  when  putting  in  gas- 
fired  house-heating  equipment,  if  he  knows  that  he  can  burn  coal  at  any 
time  he  may  desire.  For  steam  or  vapor  installations,  it  is  desirable  to 
have  the  water  line  in  both  boilers  at  the  same  level. 


PROBLEMS  IN  PRACTICE 

1  •  What  functions  must  an  automatic  stoker  perform  in  burning  coal? 

An  automatic  stoker  must  distribute  the  coal  evenly  over  the  fuel  bed,  and  fire  it  uni- 
formly. It  must  introduce  air  in  proper  quantities  to  all  parts  of  the  fuel  bed,  and  dispose 
of  the  ash  without  interfering  with  the  combustion  process.  Indirectly,  the  stoker  is 
responsible  for  the  proper  burning  of  the  volatile  fuel  gases  in  the  combustion  space 
above  the  fuel  bed. 

2  •  Classify  stokers  as  to  construction  and  operation. 

a.  Overfeed  flat  grate. 

b.  Overfeed  inclined  grate. 

c.  Underfeed  side  cleaning  type. 

d.  Underfeed  rear  cleaning  type. 

3  •  What  classification  may  be  made  of  stokers  as  to  their  use? 

Class  1.  For  residences  (Capacity  less  than  100  Ib  of  coal  per  hour). 

Class  2.  For  apartment  houses  and  small  commercial  heating  jobs  (Capacity  100  to  200 
Ib  of  coal  per  hour). 

Class  3.  For  general  commercial  heating  and  small  high  pressure  steam  plants  (Capacity 
200  to  300  Ib  of  coal  per  Hour)". 

Class  4.  For  large  commercial  and  high  pressure  steam  plants  (Capacity  over  300  Ib  of 
coal  per  hour). 

4  •  What  main  parts  are  found  in  an  underfeed  residential  stoker? 

A  hopper  is  supplied  to  hold  coal  which  is  fed  by  a  screw  >or  plunger  into  a  retort  provided 
with  air  openings  called  tuyeres.  A  Uower  supplies  air  under  pressure  for  combustion,  and 
a  gear  case  provides  for  changes  in  coal  feeding  rates. 

5  •  What  is  a  dead-plate? 

A  dead-plate  is  a  flat  surface  without  air  supply  openings  upon  which  the  fuel  rests  while 
combustion  of  the  fixed  carbon  is  completed.  Generally  the  ash  is  removed  from  the 
dead-plate.  '  v  » 

6  •  What  rate  of  coal  burning  is  usually  recommended  for  small  underfeed 
stokers? 

For  continuous  operation,  25  Ib  per  square  foot  of  grate  surface  is  recommended;  for 
short  duration  peaks,  30  Ib. 

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CHAPTER  28 — AUTOMATIC  FUEL  BURNING  EQUIPMENT 

7  •  What  methods  of  oil  atomization  are  used? 

1.  Throwing  the  oil  from  a  rotating  cup  or  disc. 

2.  Forcing  the  oil  under  high  pressure  through  a  whirl  chamber  in  a  nozzle. 

3.  Propelling  the  oil  with  a  high  velocity  jet  of  air  or  steam. 

4.  Forcing  an  oil  and  air  mixture  through  a  nozzle. 

8  •  What  is  the  purpose  of  atomization? 

Atomization  is  used  to  increase  the  surface  area  of  the  oil  in  order  to  facilitate  putting  it 
into  a  vaporous  state  so  it  may  burn. 

9  •  Is  the  furnace  of  much  importance  in  oil  burning? 

In  most  cases  it  is  very  important.  It  is  the  function  of  the  oil  burner  to  supply  the  air 
and  fuel  in  correct  proportions ;  the  furnace  must  provide  heated  space  for  proper  mixing 
and  combustion. 

10  •  Which  flame  is  considered  better,  the  luminous  or  the  non-luminous? 

Laboratory  tests  show  that  they  are  equally  efficient  in  the  usual  installation. 

11  •  What  main  precaution  is  necessary  in  choosing  a  boiler  for  an  oil  burner? 

Since  the  burner  output  is  usually  varied  through  a  wide  range  under  control  of  the 
thermostat,  a  boiler  should  be  provided  with  enough  indirect  heating  surface  to  absorb 
the  heat  as  it  is  released.  The  combustion  space  must  be  large  enough,  and  have  correct 
proportions  for  mixing  fuel  and  air  at  high  temperatures.  If  oil  is  used  inefficiently 
high  heating  costs  will  result. 

12  •  How  should  oil  burner  adjustments  be  made? 

Adjustments  should  be  made  by  an  experienced  man  who  uses  a  gas  analysis  apparatus 
to  determine  the  CO-i  content. 

13  •  What  CO2  content  should  be  attained  in  oil  burning? 

Ten  per  cent  COz  is  considered  good  practice,  for  it  indicates  the  supplying  of  50  per  cent 
excess  air. 

14  •  What  maximum  heat  release  is  considered  good  practice  in  oil  burning? 

"A  heat  release  of  38,000  Btu  per  cubic  foot  per  hour  is  considered  to  be  the  maximum  for 
average  large  installations.  This  figure  has  been  greatly  exceeded  in  some  cases,  ^  The 
design  of  the  combustion  chamber,  as  to  impingement  of  flame  and  as  to  proper  mixing 
at  high  temperatures,  has  much  to  do  with  the  attainable  heat  release. 

15  •  Name  five  types  of  gas-fired  space  heaters. 

a.  Parlor  furnaces  or  circulators. 

b.  Radiant  heaters. 

c.  Gas-fired  steam  or  hot  water  radiators. 

d.  Warm  air  radiators. 

e.  Garage  heaters. 

16  •  How  are  gas  heating  units  rated? 

Gas-fired  units  are  rated  on  the  basis  of  output  in  Btu  per  hour. 

17  •  What  safety  consideration  is  noted  in  establishing  the  ratings  of  gas-fired, 
units? 

The  rating  is  limited  by  the  amount  of  gas  that  can  be  burned  without  the  liberation  of 
harmful  amounts  of  carbon  monoxide, 

477 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

18  •  What  control  equipment  is  essential  in  the  usual  oil  burner  installation? 

See  Chapter  14.  Usually  a  room  thermostat,  a  limiting  device  to  prevent  the  pressure  or 
temperature  from  exceeding  a  desirable  limit,  a  shut-off  device  to  guard  against  failure 
of  flame  ignition,  and  in  steam  or  hot  water  boilers  a  low  water  protective  device. 

19  •  In  the  construction  of  gas-fired  garage  space  heaters,  what  special  pre- 
caution must  be  taken? 

A  safety  screen  must  be  placed  over  each  opening  into  the  combustion  chamber  to  pre- 
vent explosions  of  any  possible  gasoline  vapors. 

20  •  List  some  factors  which  might  account  for  possible  economies  of  stoker 
firing  over  hand  firing. 

a.  The  regular  feed  of  coal  instead  of  the  intermittent  feed. 
&.  The  use  of  cheaper  sizes  and  grades  of  fuel. 

c.  The  absence  of  door  openings  for  firing  purposes,  which  avoids  the  admission  of  cold 
excess  air. 

d.  The  avoidance  of  overheating  because  the  stoker  responds  quickly  to  automatic 
equipment  controlled  by  the  heat  demand. 


478 


Chapter  29 

FUEL  UTILIZATION 

Heat   Loss,    Calorific    Values,   Heating  Efficiencies,    Non-Heating 

Periods,  Heat  Capacity  oj  Buildings,  Miscellaneous  Factors,  Degree- 

Day  Method,  Rough  Approximations,  Relative  Heating  Costs 

TO  predict  the  amount  of  fuel  likely  to  be  consumed  in  heating  a 
building  during  a  normal  heating  season,  it  is  necessary  to  know  the 
total  heat  requirements  of  the  building  and  the  utilization  factor  of  the 
fuel.  The  accuracy  of  the  estimate  will  depend  on  the  ability  to  select 
these  values  and  on  the  care  taken  in  making  allowances  for  other  variable 
factors. 

Fuel  requirements1  are  given  by  the  following  general  equation  : 

H  X  (*  -  fa)  X  N  m 

(t  -  /0)  X  C  X  E  v  ' 

where 

F  =  quantity  of  fuel  required  for  a  heating  season. 
N  =  number  of  hours  of  heating  season. 
t  —  inside  temperature,  degrees  Fahrenheit. 
^  =  average  outside  temperature,  degrees  Fahrenheit, 
*o  =  outside  design  temperature,  degrees  Fahrenheit. 

H  —  calculated  heat  loss  of  building  based  on  outside  temperature  of  £Q,  Btu  per  hour. 
C  =  calorific  value  of  one  unit  of  fuel,  the  unit  being  the  same  as  that  on  which  F 

is  based. 
E  —  efficiency  of  utilization  of  fuel,  per  cent. 

HEAT  LOSS 

The  hourly  heat  loss  (H)  is  equal  to  the  sum  of  the  transmission  losses 
(fit)  and  the  infiltration  losses  (Hi)  of  the  rooms  or  spaces  to  be  heated, 

and  the  total  equivalent  heating  surface  required  is  equal  to         sq  ft. 


In  estimating  the  fuel  consumption  of  a  building  of  more  than  one  room 
divided  by  walls  or  partitions,  it  is  not  correct  to  use  the  calculated  heat 
loss  of  the  building  without  making  the  proper  allowance  for  the  fact  that 
the  heating  load  at  any  time  does  not  involve  the  sum  of  the  infiltration 
losses  of  all  of  the  heated  spaces  of  the  building  but  only  part  of  the 
infiltration  losses.  This  is  explained  in  Chapter  6. 

It  is  sufficiently  accurate  in  most  cases  to  consider  only  half  of  ^the  total 
infiltration  losses  of  a  building  having  interior  walls  and  partitions,  and 
the  value  of  H  in  Equation  1  would,  under  these  conditions,  be  equal  to 

^or  further  information  on  this  subject  see  Estimating  Fuel  Consumption,  by  Paul  D.  Close,  (Heating 
Piping  o.nd  Air  Conditioning,  May,  1931). 

479 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

Ht  +  -7T1 .    If  a  building  has  no  interior  walls  or  partitions,  whatever  air 
2> 

enters  through  the  cracks  on  the  windward  side  must  leave  through  the 
cracks  on  the  leeward  side,  and  only  half  of  the  total  crack  should  be  used 
in  computing  the  infiltration  for  each  side  and  end  of  the  building.  Under 
these  conditions  it  is  sufficiently  accurate  to  use  the  total  calculated  heat 
loss  (IT)  for  the  building.  If  the  average  wind  velocity  during  the  heating 
season  differs  from  that  upon  which  Hi  was  derived,  the  value  of  H  should 
be  corrected  accordingly. 

Of  course,  where  the  required  heating  surface  is  estimated  by  empirical 
or  rule-of-thumb  methods,  refinements  in  approximating  fuel  consump- 
tion are  not  warranted,  but  rule-of-thumb  methods  often  lead  to  unsatis- 
factory results  and  should  be  avoided  in  heating  work  where  more  accurate 
methods  are  available.  It  should  be  emphasized  that  the  value  of  H  in 
Equation  1  is  the  total  heat  loss  of  the  building  after  making  the  proper 
allowance  for  infiltration. 

CALORIFIC  VALUES  AND  HEATING  EFFICIENCIES 

The  calorific  values  of  fuel  oils  and  gas  can  be  ascertained  with  reason- 
able accuracy.  The  values  for  various  grades  of  oil  are  given  in  Table  3, 
Chapter  27.  The  calorific  value  of  gas  can  always  be  obtained  from  the 
local  utility  company.  Values  for  natural  gas  are  given  in  Table  4, 
Chapter  27;  manufactured  gas  usually  has  a  calorific  value  of  about  535. 
Coals  have  a  larger  range  and  may  vary  for  the  same  type  of  coal,  depend- 
ing on  its  ash  content.  For  general  purposes  where  specific  data  are 
lacking,  values  can  be  taken  from  Table  1,  Chapter  27. 

To  decide  on  the  correct  efficiency  to  use  is  a  more  difficult  matter,  par- 
ticularly if  the  estimate  is  being  made  without  a  full  knowledge  of  the 
equipment  for  burning  the  fuel  and  the  care  the  furnace  will  receive. 
Efficiencies  usually  are  given  in  the  catalogs  of  manufacturers  of  furnaces 
and  boilers,  but  these  values  are  obtained  under  test  conditions  and  do  not 
allow  for  poor  attendance,  defects  in  installation,  or  poor  draft.  On  the 
other  hand,  such  efficiencies  assume  that  all  the  heat  radiated  from  the 
outside  of  the  heaters  or  casings  as  sensible  heat  of  the  flue  gases  is  lost, 
whereas,  if  the  heater  is  installed  in  the  building  being  heated,  a  con- 
siderable portion  of  these  losses  may  help  to  heat  the  building2;  how  much 
of  this  it  is  legitimate  to  use  in  increasing  the  value  of  E  will  depend  on 
whether  H  included  the  heat  losses  in  the  cellar,  and  on  the  construction 
of  the  chimney.  Except  for  an  interior  chimney,  the  heat  transferred 
through  the  chimney  wall  to  the  building  will  be  very  small.  Chimney 
allowances  should  be  greater  for  lower  test  efficiencies.  Thus  an  insulated 
furnace  will  give  a  high  efficiency  on  test  but  will  not  heat  the  cellar.  A 
modern  gas  furnace  will  have  a  high  efficiency  with  a  correspondingly  low 
flue  gas  temperature  and  hence  there  will  be  very  little  heat  from  the 
flue  pipe. 

For  great  exactitude  the  value  for  E  should  take  care  of  inefficiency  in 
the  heat  distribution  in  the  building  because  of  such  losses  as  excessive 
heating  of  the  walls  behind  the  radiators  and  excessive  stratification.  It 

'Analysis  of  the  Over-All  Efficiency  of  a  Residence  Heated  by  Warm  Air,  by  A.  P.  Kratz  and  J.  F. 
Quereau  (A.S.H.V.E.  TRANSACTIONS,  Vol.  35,  1929). 

480 


CHAPTER  29 — FUEL  UTILIZATION 


is  preferable,  however,  to  include  these  losses  in  the  value  of  H,  and  to 
limit  E  to  the  fuel  burning  equipment. 

Automatic  fuel  burning  equipment,  whether  for  coal,  oil  or  gas,  will  tend 
to  save  fuel  and  will  therefore  produce  a  higher  efficiency  if  thermostati- 
cally controlled,  but  on  the  other  hand  automatic  equipment  tends  to 
make  the  householder  prolong  his  heating  season  and  maintain  a  higher 
temperature  in  the  house  in  the  early  fall  and  late  spring. 

NON-HEATING  PERIODS 

Obviously,  the  theoretical  fuel  consumption  will  be  reduced  con- 
siderably by  not  operating  the  heating  plant  at  night.  Allowance  for 
this  may  be  made  in  either  of  two  ways :  (1)  by  estimating  the  average 
inside  temperature  t,  or  (2)  by  arbitrarily  assuming  a  certain  reduction 
in  the  fuel  consumption. 

The  first  procedure  is,  of  course,  the  more  accurate.  If,  for  example,  the 
daytime  temperature  is  to  be  70  F,  and  the  temperature  from  12  midnight 
to  6  a.m.  is  to  be  maintained  by  thermostatic  control  'at  50  F,  then  the 

average  daily  inside  temperature  t  will  be ~- or  65  F. 

Strictly  speaking,  this  average  inside  temperature  would  apply  only  when 
the  outside  night  temperature  averages  below  50  F,  but  this -fact  usually 
is  not  of  sufficient  importance  to  warrant  consideration.  If  the  average 
outside  temperature  during  the  heating  season  is  30  F,  the  fuel  saving 

f-rr\    £>  pr 

would  be  approximately  100  X  ^ ™  or  12.5  per  cent.    In  this  case,  the 

/U  —  oU 

additional  saving  in  fuel  due  to  the  cooling  of  the  air  and  structural 
materials  to  50  F  would  be  offset  by  the  heating-up  load  in  the  morning. 
As  to  the  second  procedure,  it  may  be  arbitrarily  assumed  that  a 
saving  in  the  fuel  consumption  of  from  10  to  30  per  cent,  depending  on 
conditions,  will  result  if  the  heat  is  shut  off  after  working  hours,  and  the 
building  is  heated  to  the  required  temperature  during  the  period  of  occu- 
pancy each  day.  This,  of  course,  is  a  general  statement  and  wherever 
possible  the  average  temperature  should  be  estimated  from  the  propor- 
tionate lengths  of  the  occupancy  and  non-occupancy  periods  and  the  cor- 
responding temperatures  for  these  periods.  Any  deviation  from  the  assumed 
inside  temperature  will  result  in  a  variation  in  the  estimated  fuel  con- 
sumption. 

HEAT  CAPACITY  OF  BUILDINGS 

The  heat  required  to  warm  the  cold  building  and  contents  is  a  factor 
to  be  considered.  Under  certain  conditions,  the  cooling  of  the  structure 
and  contents  will,  to  some  extent,  compensate  for  the  heat  required  to 
rewarm  the  building.  For  example,  if  the  building  is  under  thermostatic 
control  and  the  day  and  night  temperatures  are  say  70  F  and  50  F, 
respectively,  there  will  be  a,  period  during  which  no  heat  will  be  called  for 
while  the  building  is  cooling  to  50  F,  and  the  saving  resulting  therefrom  will 
correspond  to  the  additional  heat  required  to  bring  the  building  and  con- 
tents back  to  the  daytime  temperature.  If  in  estimating  the  fuel  con- 
sumption the  average  daily  inside  temperature  is  based  on  the  proper  day 
and  night  temperatures  and  periods,  the  heat  required  to  warm  the 
structure  may  be  neglected. 

481 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

Where  irregular  conditions  are  involved  it  may  be  desirable  to  actually 
calculate  the  fuel  required  to  warm  the  building  structure  and  contents 
for  the  number  of  times  during  the  heating  season  the  heating  plant  would 
not  be  in  operation  and  to  add  this  quantity  to  the  fuel  required  for  the 
number  of  hours  during  which  the  building  is  heated.  The  greater  the 
heat  capacity  of  the  structure  the  greater  will  be  the  relative  importance 
of  this  item.  For  structures  of  low  heat  capacity,  such  as  frame  buildings, 
this  factor  usually  may  be  neglected. 

Example  1 .  A  small  factory  building  located  in  Philadelphia  is  to  be  heated  to  60  F 
between  the  hours  of  7  a.m.  and  7  p.m.,  and  to  50  F  during  the  remaining  hours.  The 
calculated  hourly  heat  loss,  based  on  a  design  temperature  of  —  6  F,  is  500,000  Btu.  If 
coal  having  a  calorific  value  of  12,500  Btu  per  pound  is  fired,  and  the  over-all  heating 
efficiency  is  assumed  to  be  60  per  cent,  how  many  tons  of  coal  will  be  required  for  a 
normal  heating  season,  neglecting  other  heat  sources  and  any  loss  of  heat  through  open 
windows? 

Solution.  Since  there  are  no  partitions  in  this  building,  the  entire  heat  loss  is  con- 
sidered. The  average  outside  temperature  during  the  heating  season  Za  is  41.9  F  (see 
Table  2,  Chapter  7);  t  =  60  F;  N  «  5040;  H  =  500,000;  (t  -  /0)  =  66  F;  C  =  12,500; 
E  =  0.60.  Substituting  these  values  in  Equation  1  and  dividing  by  2000  to  change  to 
tons: 

„          500,000  X  18.1  X  5040  -        . 

F  ~  66X12,500X0.60X2000  =  46  t0nS  °f  C°al 

Inasmuch  as  the  building  will  be  heated  to  50  F  at  night,  the  average  inside  tempera- 
ture at  the  breathing  line  will  be  55  F,  and  the  percentage  saving  will  be  QQ  _  A-\  Q 

=  0.276  or  27.6  per  cent.    The  net  fuel  consumption  will  therefore  be  46  -  0.276  X  46 
or  33.3  tons. 

MISCELLANEOUS  FACTORS 

There  are  many  factors  which  would  be  likely  to  affect  the  theoretical 
fuel  requirements  of  a  building,  such  as  the  opening  of  windows,  abnormal' 
inside  temperatures,  other  heat  sources,  sun  effect,  wind,  and  rain.  In 
many  cases  it  is  difficult  to  evaluate  these  factors  accurately,  particularly 
in  the  case  of  open  windows,  and  the  results  are  correspondingly  less 
accurate.  The  degree  of  refinement  of  the  calculations  should,  of  course, 
be  consistent  with  the  conditions  involved.  If  the  heat  loss  from  the 
boiler  and  piping  does  not  warm  the  building  or  is  not  included  in  £T, 
the  proper  allowance  should  be  made.  In  selecting  a  boiler,  this  allowance 
is  frequently  assumed  to  be  25  per  cent  of  the  total  heat  loss  of  the  build- 
ing, but  in  estimating  fuel  requirements,  the  more  accurate  procedure  of 
computing  the  pipe  and  boiler  losses  should  be  used,  unless  this  item  is 
likely  to  be  outweighed  by  other  less  tangible  factors. 

Where  temperature  control  is  installed  the  fuel  consumption  can 
obviously  be  predetermined  with  greater  accuracy  than  where  no  such 
control  has  been  provided.  In  fact  the  calculated  requirements  agree  to  a 
remarkable  extent  in  many  cases  with  the  actual  fuel  consumption.  This 
has  been  particularly  true  of  gas-fired  installations,  with  which  effective 
temperature  regulation  usually  is  possible. 

OTHER  HEAT  SOURCES 

Where  other  heat  sources  are  available  it  is  quite  often  possible  to  make 
accurate  allowance  for  the  reduction  in  the  fuel  consumption  resulting 

482 


CHAPTER  29 — FUEL  UTILIZATION 


therefrom.  These  sources  include  the  heat  supplied  by  persons,  lights, 
motors  and  machinery,  and  should  also  be  ascertained  in  the  case  of 
theaters,  assembly  halls  and  industrial  plants.  (See  Chapter  7.)  In  many 
cases  these  heat  sources  should  not  be  allowed  to  affect  the  size  of  the  in- 
stallation of  heating  equipment,  although  they  may  have  a  marked  effect 
upon  the  fuel  consumption.  In  residences  this  factor  usually  may  be 
neglected. 

DEGREE-DAY  METHOD 

A  very  useful  unit  for  estimating  fuel  consumption,  particularly  for 
residences,  is  the  degree-day.  (See  definition  in  Chapter  41.)  Degree-days 
for  various  cities  in  the  United  States  and  Canada  are  given  in  Table  1. 
The  term  degree-day  originated  in  the  gas  industry  and  was  later  stand- 
ardized by  the  American  Gas  Association*. 

The  base  of  65  F  is  used  for  an  inside  temperature  of  70  F.  This  base 
was  chosen  because  it  was  demonstrated,  by  means  of  data  collected  from 
numerous  installations,  that  heat  is  seldom  supplied  to  a  residence  when 
the  outdoor  temperature  is  greater  than  65  F.  It  was  also  found  that 
the  fuel  consumed  varied  almost  directly  with  the  difference  between 
65  F  and  the  outside  temperature. 

If  the  inside  temperature  were  maintained  at  70  F  throughout  the  24 
hours  of  the  day,  then  the  base  of  65  F  would  probably  be  in  error.  It 
must  be  borne  in  mind,  however,  that  although  the  temperature  head  is 
the  difference  between  the  inside  temperature  of  say  70  F,  and  the  outside 
temperature,  a  lower  temperature  than  70  F  will  usually  be  maintained  at 
night  and  the  base  of  65  F  will  therefore  allow  for  this  condition.  As 
already  indicated,  a  temperature  of  50  F  from  midnight  to  6  a.m.  will 
reduce  the  24-hour  average  from  70  to  65  F.  It  is  important  to  note  that 
the  degree-day  applies  specifically  to  an  inside  temperature  of  70  F, 
which  is  the  usual  temperature  for  residences,  and  it  should  also  be  noted 
that  allowance  is  automatically  made  for  the  lower  nighttime  tempera- 
ture, although  this  allowance  is  constant  for  any  given  locality. 

In  Equation  1,  the  quantity  (t  —  4)  X  ^V  is  equivalent  to  the  number 
of  degree-days  D  in  a  heating  season  multiplied  by  24,  when  the  average 
daily  value  of  t  is  65  F.  Therefore, 

(t  -  /a)  X  N  -  24  D  (2) 

Substituting  the  value  of  (t  —  4)  X  N  from  Equation  2  in  Equation  1, 
the  following  general  formula  for  an  average  daily  inside  temperature  of 
65  F,  which  is  approximately  equivalent  to  an  inside  daytime  temperature 
of  70  F  for  residences,  is  obtained: 

Fd  "  '(t  -  fc)  X  C  X  E  (3) 

Example  2.  The  calculated  hourly  heat  loss  of  a  residence  located  in  Chicago  is 
127,000  Btu,  which  includes  28,000  Btu  for  infiltration.  The  design  temperatures  are 
—  8  If  and  70  F.  The  normal  heating  season  is  assumed  to  be  210  days  (5,040  hours)  and 
the  average  temperature  during  this  period  is  36.4  F  (see  Table  2,  Chapter  7).  The 


*See  Industrial  Gas  Series,  House  Heating  (third  edition)  published  by  the  American  Gas  Association, 

483 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


TABLE  1.    DEGREE-DAYS  FOR  CITIES  IN  THE  UNITED  STATES  AND  CANADAa 


Col.  A 

ColB 

ColC 

Col  A 

ColB 

Col.  C 

State 

City 

Degree-Days 

State 

City 

Degree-Days 

Ala. 

Birmingham 

2,408 

Nev.      . 

Reno  

5,891 

Mobile 

1  471 

N  H 

Concord  

6,852 

Ariz 

Flagstaff 

7,145 

N.  T 

Atlantic  City  

5,175 

Tucson 

1  845 

Trenton  

4,934 

Ark 

Hot  Springs 

2,665 

N  M. 

Santa  Fe  

6,063 

Little  Rock 

2811 

N  Y 

Albany  

6,889 

Calif. 

Los  Angeles 

1,504 

Buffalo  

6,821 

San  Francisco 

3264 

New  York  ...        

5,348 

Colo 

Colorado  Springs 

6553 

N.  C. 

Raleigh  

3,234 

Denver 

5  873 

Wilmington  

2,302 

Conn 

New  Haven 

5  895 

N   Dak 

Bismarck 

8,498   ' 

D  C 

Washington 

4626 

Ohio 

Cincinnati  

4,702 

Fla 

Jacksonville 

890 

Cleveland  

6,154 

Ga 

Atlanta             .... 

2,891 

Columbus  

5,323 

Savannah 

1,490 

Okla. 

Oklahoma  City  

3,613 

Idaho 

Boise  

4,558 

Ore  

Portland  

4,468  ' 

Lewiston 

4924 

Salem  

4,629 

111. 

Chicago  

6,315 

Pa  

Philadelphia  

4,855 

Springfield 

5,370 

Pittsburgh  . 

5,235 

Ind 

Evansville 

4  164 

R  I. 

Providence 

6,014 

Indianapolis 

5,297 

S.  C.. 

Charleston  .•. 

1,769 

Iowa 

Des  Moines  

6,373 

Spartanburg.  

3,257 

Sioux  City 

7,023 

S.  Dak. 

Sioux  Falls  

7,683 

Kans. 

Dodge  City     

5,034 

Tenn  

Memphis  

2,950 

Topeka  . 

5,301 

Nashville  

3,578 

Kv 

Lexington 

4616 

Texas 

Austin 

1,578 

Louisville 

4,180 

Dallas  

2,455 

La 

New  Orleans 

1,023 

Houston 

1,157 

Me. 

Eastport 

8,531 

San  Antonio  . 

1,202 

Portland 

7,012 

Utah 

Logan 

6,735 

Md. 

Baltimore 

4,333 

Salt  Lake  City  

5,553 

Mass. 

Springfield 

6,464 

Vt.  . 

Burlington 

7,620 

Boston 

6,145 

Va.  .. 

Fredericksburg 

4,243 

Mich 

Detroit 

6494 

Norfolk 

3,349 

Marquette    . 

8,692 

Richmond  

3,725 

Minn  

Duluth  

9,480 

Wash  

Seattle  

4,868 

Minneapolis 

7,851 

Spokane  

6,353 

Miss  

Vicksburg  

1,822 

W.  Va  

Morgantown  

5,016 

Mo.    . 

Kansas  City 

5,202 

Parkersburg 

4,884 

St.  Louis  

4,585 

Wis  

Fond  du  Lac  

7,612 

Mont  

Billings. 

7,115 

Green  Bay.  .  . 

7,823 

Havre  

8,699 

La  Crosse  

6,690 

Nebr  

Lincoln. 

6,231 

Milwaukee 

7,372 

Omaha  

6,128 

Wyo.  .      . 

Cheyenne  

7,462 

Province 

City 

Degree-Days 

Province 

City 

Degree-Days 

B.  C 

Victoria 

5,777 

Ont. 

Toronto 

7,732 

Vancouver  

5,976 

Que  

Montreal  

8,705 

Kamloops   . 

6,724 

Quebec 

8,628 

Alb. 

Medicine  Hat 

8,152 

N.  B. 

Fredericton 

9099 

Sask. 

Ou'AoDelle 

11,261 

N.  S.  . 

Yarmouth 

7,694 

Man.  ....   . 

Winnipeg  

11,166 

P.E.I.  ... 

Charlottetown  

8,485 

Ont 

Port  Arthur 

10,803 

aFrom  Industrial  Gas  Aeries,  House  Heating  (third  edition)  published  by  the  American,  Gas  Association. 
These  degree-days  are  based  on  daily 'mean  temperatures.    Base,  65  F. 

484 


CHAPTER  29 — FUEL  UTILIZATION 


building  is  to  be  heated  with  oil  fuel  having  a  calorific  value  of  141,000  Btu  per  gallon. 
The  heating  efficiency  is  assumed  to  be  70  per  cent.  Thermostatic  control  is  to  be  used 
and  a  temperature  of  55  F  is  to  be  maintained  from  11  p.m.  to  7  a.m.  How  many  gallons 
of  oil  will  'be  required  during  a  normal  heating  season  if  the  loss  of  heat  through  open 
windows  is  neglected? 


Solution.    The  maximum  hourly  heat  loss  will  be  127,000  -  =  113,000  Btu 

2t 
=  H.    Substituting  the  proper  values  in  Equation  1: 

113,000  X  (70  -  36.4)  X  5040 
F  =  141,000  X  0.70  X  [70  -  (-  8)1  =  2486  gal  °f  Olh 

T«  .     .  ,    ,  .„  ,70  X  16  +  55  X  8       _  « 

The  average  inside  temperature  will  be  -  ^7  -  =  65  F 

and  the  fuel  saving  due  to  this  fact  will  be   ^  -  5—  =  0.149  or  14.9  per  cent. 

— 


Hence,  the  net  fuel  consumption  will  be  2486  —  0.149  X  2486  =  2116  gal.      - 
The  normal  number  of  degree-days  for  Chicago  is  6315.    Substituting  in  Equation  3 
and  solving  by  the  degree-day  method: 

_,       113,000  X  6315  X  24        OOOK      .    ,    .. 
F  =    78  X  141,000  X  0.70    =  2225  gal  °f  Ol1 

No  allowance  need  be  made  for  the  average  temperature  of  65  F  since  this  is  taken 
care  of  by  the  selection  of  a  base  of  65  F  for  the  degree-day,  as  already  explained.  It  will 
be  noted  that  the  two  methods  check  within  5  per  cent  in  this  case.  If  the  average  daily 
inside  temperature  in  the  first  solution  had  been  66.4  F  instead  of  65  F,  the  two  methods 
would  have  checked  exactly. 

INDUSTRIAL  DEGREE-DAY 

Since  the  standard  degree-day  is  intended  for  an  inside  temperature  of 
70  F,  it  is  particularly  convenient  for  solving  residence  problems.  Where 
the  design  temperature  differs  greatly  from  70  F,  the  standard  degree-day 
cannot  be  accurately  applied*  Consequently,  the  industrial  degree-day4 
has  been  developed  and  values  have  been  derived  for  two  bases,  namely 
55  F  and  45  F,  intended  for  inside  temperatures  of  60  F  and.  50  F,  re- 
spectively. 

There  is  a  considerable  spread,  however,  among  these  three  bases,  and 
consequently  there  would  be  an  appreciable  error  if  the  actual  basis  to  be 
used  in  a  certain  case  would  be  approximately  midway  between  any  two 
of  the  three  bases  for  which  degree-day  values  are  at  present  available. 
Since  the  correction  cannot  be  made  on  a  proportionate  basis,  it  would  be 
more  accurate  in  the  majority  of  cases  involving  inside  temperatures  other 
than  70  F,  60  F,  or  50  F  to  apply  Equation  1. 

APPROXIMATING  FUEL  REQUIREMENTS 

It  is  sometimes  desirable  to  obtain  a  rough  approximation  of  the  annual 
fuel  consumption.  Such  approximations  may  be  obtained  by  using  unit 
factors  based  on  the  fuel  requirements  per  square  foot  (or  per  100  sq  ft)  of 
radiation  or  per  1000  cu  ft  of  space. 

Fig.  1  may  be  used  for  rough  approximations  of  coal  and  oil  require- 


4See  Heating  and  Ventilating  Degree- Day  Handbook, 

485 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


ments.  It  should  be  noted  that  this  figure  is  given  in  terms  of  the  fuel 
consumption  per  1000  degree-days  per  100  sq  ft  of  equivalent  heating 
surface  (steam)  based  on  an  emission  of  240  Btu  per  square  foot.  Unless 
the  radiation  is  calculated  with  reasonable  accuracy,  unit  factors  will  be  of 
little  value  even  for  rough  approximations,  since  it  is  obvious  that  such 
radiation  requirements  must  bear  some  relationship  to  the  actual  heating 
requirements  of  the  building. 

Example  3.  Estimate  the  approximate  coal  consumption  for  a  building  located  in 
New  York  City  in  which  the  calculated  heating  surface  requirements  (steam)  are 
1000  sq  ft  based  on  design  temperatures  of  zero  and  70  F. 


Tons  of  Coal  per  1000  Defr-Days  per  100  Sq.  Ft  of  Heating  Surface, 
240  B.  t.  u.  per  sq.  ft 
D  p  p  p  a 

5  g  S  S  8 

- 

/ 

s  §  s  g  s  s 

Gals,  of  Oil  per  1000  Deg.-Days  per  100  Sq  Ft  of  Heatmg  Surface, 
240  B.  t.  u.  per  sq  ft 

/ 

- 

/ 

• 

- 

/ 

/ 

- 

y  , 

/ 

- 

/ 

~" 

/ 

/ 

I  

-^ 

/ 

-10 


0,  +10 

Outside  Design  Temp.,  deg.  fahr. 


+20 


FIG.  1.   CURVE  FOR  OBTAINING  ROUGH  APPROXIMATION  OF  ANNUAL  FUEL  CONSUMPTION 

IN  TONS  OF  COAL  OR  GALLONS  OF  OIL  PER  1000  DEGREE- DAYS  PER 

100  SQ  FT  OF  EQUIVALENT  STEAM  HEATING  SURFACE** 

aThis  curve  is  based  on  heating  efficiencies  of  60  and  70  per  cent  for  coal  and  oil,  respectively,  a  calorific 
value  of  coal  of  13,000  Btu  per  pound,  a  calorific  value  of  oil  of  140,000  Btu  per  gallon,  an  inside  tempera- 
ture of  70  F,  and  an  emission  of  240  Btu  per  equivalent  square  foot  of  heating  surface  (steam),  and  does  not 
allow  for  unusual  factors  which  would  affect  the  fuel  consumption,  such  as  open  windows,  week-end  shut- 
downs, etc.  For  hot  water,  divide  the  result  obtained  by  means  of  this  chart  by  1.6. 

Solution,  From  Fig.  1,  the  fuel  consumption  for  a  design  temperature  of  zero  is  0.53 
tons  per  1000  degree-days  per  100  sq  ft  of  heating  surface.  Since  there  are  5348  degree- 
days  in  New  York  City  in  a  normal  heating  season,  the  fuel  consumption  will  be  approxi- 
mately 0.53  X  5.348  X  10  =  28.34  tons. 

Fig.  2  is  taken  from  the  3rd  edition  of  Industrial  Gas  Series  on  House 
Heating,  published  by  the  American  Gas  Association,  and  indicates  the 
average  gas  consumption  per  degree-day  for  various  heat  contents. 
While  the  fuel  consumption  in  individual  cases  may  vary  somewhat  from 
the  curve  values,  these  average  values  are  sufficiently  accurate  for  esti- 
mating purposes  and  give  very  satisfactory  results. 

The  value  generally  used  in  the  manufactured  gas  industry  for  resi- 
dences is  0.21  cu  ft  per  degree-day  per  square  foot  of  equivalent  steam 

486 


CHAPTER  29 — FUEL  UTILIZATION 


radiation  (240  Btu)  based  on  the  theoretical  requirements.  A  correction 
for  warmer  climates  is  necessary  and  it  is  customary  to  gradually  increase 
the  relative  fuel  consumption  below  3,000  degree-days  to  about  20  per 
cent  more  at  1,000  degree-days. 

For  hot  water  or  warm  air  heat  the  fuel  consumption  is  about  0.19  cu  ft 
per  degree-day  per  square  foot  of  equivalent  steam  radiation,  that  is,  per 
240  Btu  per  hour.  The  actual  requirements  likewise  relatively  increase 
with  hot  water  or  warm  air  systems  as  the  number  of  degree-days  de- 
creases below  3,000.  For  larger  installations,  that  is,  1,000  sq  ft  of 


D 

o: 


I 


200      300 


r-cu-'m  OF  CAS  PER  SQ.  FT.  INSTALLED 

OT    WATER     RADIATION  , 


IHOT 


2-CU  FT.   OF    GAS    PER  5<5>.  FT.    INSTALLED 
i         i         i         STEAM      RAD.IATION     , 


3  -CU.  FT.  OF    GAS    PER    100   CU>  FT.  BLDC. 
|         |  CONTENTS    HOT.  AIR. SYSTEM.         , 


4-CU.FT.  OF    GAS  PER    1000   B.T.U.   HOURLY 
LOSS   FROM     BLDG-  HOT    AIR    SYSTEM 


400       500      600      700       800 
B.T.U.   VALUE   OF   GAS 


900       1000 


FIG.  2.    CHART  GIVING  GAS  REQUIREMENTS  PER  DEGREE-DAY  FOR  VARIOUS  CALORIFIC 
VALUES  OF  GAS  AND  FOR  DIFFERENT  HEATING  SYSTEMS* 

aThis  chart  is  based  on  an  inside  temperature  of  70  F  and  an  outside  temperature  of  zero.  If  the  radia- 
tion is  installed  on  the  basis  of  any  other  temperature  difference,  multiply  the  result  obtained  from  this 
chart  by  70,  and  divide  by  the  actual  temperature  difference. 

theoretical  radiation  and  above,  there  is  an  increase  in  efficiency,  and  a 
consequent  decrease  in  the  fuel  consumption  per  degree-day  per  square 
foot  of  heating  surface. 

The  approximate  quantities  of  steam  required  in  New  York  City  per 
square  foot  of  heating  surface  for  various  classes  of  buildings  are  given  in 
Chapter  37. 

The  preceding  discussion  on  fuel  consumption  has  dealt  with  the  heating 
requirements  of  the  building  irrespective  of  any  air  that  may  be  intro- 
duced for  ventilation  purposes  other  than  the  normal  infiltration  of  out- 
side air.  The  heat  required  for  warming  air  brought  into  the  building  for 
ventilation  may  be  estimated  from  data  given  in  Chapters  2  and  22. 

487 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

RELATIVE  HEATING  COSTS 

A  comparison  of  the  relative  cost  of  heating  with  different  fuels  can 
be  made  with  even  a  fair  degree  of  accuracy  only  when  there  is  a  full 
knowledge  of  the  equipment  which  will  be  used  with  each  fuel,  and  the 
efficiency  with  which  each  will  be  operated.  When  proposing  to  sub- 
stitute one  fuel  for  another,  the  yearly  cost  with  the  fuel  being  used  can  be 
obtained.  The  accuracy  of  the  comparison  will  depend  upon  the  care 
taken  in  estimating  the  cost  of  the  new  fuel  with  the  equipment  which 
will  be  used. 

A  convenient  basis  for  comparison  of  various  fuels  is  the  cost  per 
million  Btu.  The  formula  used  in  estimating  costs  for  coal  is: 

Y  -    500XC  (4) 

*  ~    CcXEc  W 

where 

X  —  cost  of  heating  with  coal  in  dollars  per  million  Btu. 
c  =  cost  of  coal  in  dollars  per  ton. 
Cc  =  calorific  value  of  coal,  Btu  per  pound. 
EC  =  over-all  or  house  efficiency  for  coal,  expressed  as  a  decimal. 

Example  4.  If  coal  having  a  calorific  value  of  13,000  Btu  per  pound  costs  $10.00  per 
ton,  the  cost  per  million  Btu,  assuming  an  efficiency  of  60  per  cent,  will  be: 

50°  X<        -    n  64 


_  _ 

~  13,000  X  0.60 

The  formula  used  in  estimating  costs  for  oil  is  : 


_ 

C0  X  W  X  £0 
where 

Y  =  cost  of  heating  with  oil  in  dollars  per  million  Btu. 
p  —  cost  of  oil  in  dollars  per  gallon. 
C0  —  calorific  value  of  oil,  Btu  per  pound, 
W  —  weight  of  oil  per  gallon,  pounds. 
Eo  -  over-all  or  house  efficiency  for  oil,  expressed  as  a  decimal. 

Example  5.  If  oil  having  a  calorific  value  of  141,000  Btu  per  gallon  (C0  X  W)^  costs 
10  ji  per  gallon,  the  cost  per  million  Btu,  assuming  an  efficiency  of  70  per  cent,  will  be: 

1,000,000x0.10  , 

y         141,000  X  0.70 
The  formula  used  in  estimating  costs  for  gas  is: 

--cr^  » 

where 

Z  =  cost  of  heating  with  gas  in  dollars  per  million  Btu. 

g  =  average  cost  of  gas,  including  demand  and  commodity  charges,  dollars  per 

thousand  cubic  feet. 

Cg  =  calorific  value  of  gas,  Btu  per  cubic  foot. 
E%  =s  over-all  or  house  efficiency  for  gas,  expressed  as  a  decimal. 

Example  6,  If  manufactured  gas,  having  a  calorific  value  of  535  Btu  per  cubic  foot, 
costs  60?  per  thousand  cubic  feet,  the  cost  per  million  Btu,  assuming  an  efficiency  of 
80  per  cent,  will  be: 


535  X'0.80 

488 


CHAPTER  29 — FUEL  UTILIZATION 


PROBLEMS  IN  PRACTICE 

1  •  What  two  factors  are  most  essential  in  estimating  the  fuel  consumption  for 
heating  a  building  in  a  normal  season? 

The  total  heat  requirements  of  the  building,  and  the  efficiency  of  combustion. 

2  •  What  will  be  the  cost  per  year  of  heating  a  building  with  gas,  assuming  that 
the  calculated  hourly  heat  loss  is  92,000  Btu  based  on  0  F,  which  includes 
26,000  Btu  for  infiltration?     The  design  temperatures  are  0  F  and  72  F.     The 
normal  heating  season  is  210  days,  and  the  average  outside  temperature  during 
the  heating  season  is  36.4  F.     The  heating  efficiency  will  be  75  per  cent.     The 
heating  plant  will  be  thermostatically  controlled,  and  a  temperature  of  55  F 
will  be  maintained  from  11  p.m.  to  7  a.m.     Assume  that  the  price  of  gas  is 
7  cents  per  100,000  Btu  of  fuel  consumption,  and  disregard  the  loss  of  heat 
through  open  windows  and  doors. 

The  maximum  hourly  heat  loss  will  be 

92,000  -  ~^^   =  79>0QO  Btu  =  H. 

2> 

_       79,000  X  (72  -  36.4)  X  24  X  210 


100,000  X  0.75  X  (72  -  0) 
The  average  inside  temperature  will  be 

(72  X  16)  +  (55  X  8) 
24 

The  fuel  saving  will  be 

100  (72  -  66.3)        ,c 


oco,  „,.,„,  ,  0, 

2624'9  hundred  th°USatld  Btu' 


Per  Cent' 


72  -  36.4 
Hence,  the  net  fuel  consumption  will  be 

2624.9  -  0.16  X  2624.9  =  2204.9  hundred  thousand  Btu. 
2204.9  X  0.07  =  $154.34  =  the  cost  per  year  of  heating  the  building. 

3  •  What  factors  should  be  taken  into  consideration  when  determining  the 
efficiency  at  which  a  fuel  will  be  burned? 

Manufacturers'  catalogs  usually  give  equipment  efficiencies  obtained  under  test  con- 
ditions. These  values  do  not  allow  for  poor  attendance,  defects  in  installation,  or  poor 
draft.  Such  efficiencies  do  not  consider  heat  radiated  from  the  outside  of  the  equipment, 
but  in  many  cases  this  heat  is  utilized. 

4  •  If  20  tons  of  coal  having  a  calorific  value  of  13,000  Btu  per  pound  are  burned 
in  a  warm  air  furnace  and  produce  286,000,000  Btu  at  the  bonnet,  what  is  the 
efficiency  of  the  furnace? 

Number  of  Btu  at  bonnet 


Number  of  tons  X  calorific  value  X  number  of  Ib  in  one  ton 
286,000,000  X  100 


efficiency. 


20  X  13,000  X  2000 


=  55  per  cent. 


5  •  In  making  degree-day  calculations,  why  is  the  base  of  65  F  used  for  an  in- 
side temperature  of  70  F? 

This  base  was  chosen  because  data  collected  from  numerous  installations  show  that  heat 
is  seldom  supplied  to  a  residence  when  the  outdoor  temperature  is  greater  than  65  F.  It 
was  also  found  that  the  amount  of  fuel  consumed  varied  in  almost  direct  proportion  with 
the  difference  between  65  F  and  the  outside  temperature. 

489 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

6  •  Make  a  rougli  approximation  of  the  amount  of  coal  required  to  heat  a 
building  located  in  Cleveland,  Ohio,  assuming  that  the  calculated  heating 
surface  requirements  are  500  sq  ft  of  steam  radiation  based  on  design  tem- 
peratures of  70  F  and  0  F. 

Using  Fig.  1,  the  fuel  consumption  for  a  design  temperature  of  0  F  is  found  to  be  0.53 
tons  per  thousand  degree-days  per  hundred  square  feet  of  heating  surface.  Cleveland 
has  a  heating  season  equivalent  to  6154  degree-days,  therefore, 

0.53  X  6.154  X  5  =  16.31  tons  of  coal. 

7  •  Make  a  rough  approximation  of  the  gas  required  to  heat  a  building  located 
in  Chicago,  111.,  assuming  that  the  calculated  heating  surface  requirements 
are  1000  sq  ft  of  hot  water  radiation  based  on  design  temperatures  of  0  F  and 
70  F.    Chicago  has  800-Btu  mixed  gas,  and  6315  degree-days. 

Using  Fig.  2,  the  fuel  consumption  for  a  design  temperature  of  0  F  with  800:Btu  gas  is 
found  to  be  0.08  cu  ft  of  gas  per  degree-day  per  square  foot  of  hot  water  radiation. 

0.08  X  6315  X  1000  =  505,200  cu  ft. 

8  •  A  certain  building  has  a  maximum  heat  loss  of  250,000  Btu  per  hour  in  —  15  F 
weather.    How  many  tons  of  fuel  will  be  required  to  maintain  a  temperature  of 
70  F  during  a  260-day  heating  season  in  which  the  average  temperature  is  39  F? 
The  heating  value  of  the  fuel  is  13,200  Btu  per  pound  and  the  efficiency  of  com- 
bustion is  60  per  cent. 

250,000  (70  -  39)  260  X  24 
(70  +  15)  13,200  X  0.60  X  2000 

9  •  Which  item  may  be  determined  more  closely,  the  heating  value  of  a  fuel  or 
the  efficiency  of  its  combustion? 

The  heating  values  of  oil,  gas,  and  solid  fuels  are  closely  determinate,  whereas  the 
efficiency  of  burning  depends  on  the  particular  equipment  chosen  and  the  skill  used  in 
handling  it. 

10  •  In  an  office  building,  the  thermostats  are  set  to  maintain  70  F  from  7  a.m. 
to  5  p.m.  and  50  F  during  the  rest  of  the  time.    When  the  outside  temperature 
is  30  F,  how  much  saving  might  be  expected  because  the  temperatures  are 
lowered?    Under  the  above  conditions  the  building  becomes  50  F  by  11  p.m.  and 
warms  up  to  70  F  by  8  a.m. 

A  temperature  of  70  F  is  maintained  during  9  hours,  and  one  of  50  F  during  8  hours;  the 
temperature  would  average  about  60  F  during  the  7  hours  required  for  cooling  down  and 
warming  up.  The  average  is  60.4  for  the  24  hours.  (The  average  temperature  calcu- 
lated would  have  been  58.3  F,  had  the  warming  and  cooling  periods  been  neglected.) 


The  saving  is  jj^r          X  100  =  -       X  100  =  24  per  cent. 

11  •  How  does  the  heat  capacity  of  a  structure  influence  the  saving  made  by 
carrying  lower  temperatures  during  the  night? 

The  heat  storage  capacity  of  the  walls  prevents  rapid  dropping  of  temperatures  at  night- 
time  and  delays  the  warming  up  process  in  the  morning.  In  an  extreme  case,  the  building 
would  not  reach  the  lowered  temperature  by  the  time  the  higher  temperature  is  called 
for  in  the  morning.  But  under  any  conditions,  the  saving  made  by  lowering  the  tem- 
perature can  be  correctly  estimated  by  using  the  average  temperature  observed  over  the 
24-hour  period  as  a  factor,  as  in  Question  10. 

12  •  What  are  some  of  the  miscellaneous  factors  that  may  cause  actual  fuel 
consumption  to  vary  from  the  theoretical  fuel  requirements  as  calculated  by 
the  use  of  heat  losses,  temperature  difference,  and  fuel  burning  efficiency? 

The  opening  of  windows;  abnormally  high  or  low  inside  temperatures;  other  sources  of 
heat,  such  as  machinery  or  lights;  sun  effect;  and  unusual  winds. 

490 


Chapter  30 

RADIATORS  AND  GRAVITY 
CONVECTORS 

Heat  Emission  of  Radiators  and  Convectors,  Types  of  Radiators, 

Output  of  Radiators,  Heating  Effect,  Heating  Up  the  Radiator, 

Enclosed    Radiators,    Convectorsy    Code    Tests,     Gravity -Indirect 

Heating  Systems 

THE  general  terms  for  heating  units  are:  (1)  radiators,  for  direct  sur- 
faces, either  exposed,  enclosed,  or  shielded;  and  (2)  connectors,  or 
concealed  heaters,  for  extended  surfaces  that  are  built  in  as  part  of  an 
enclosure  or  cabinet.     Some  heating  units  are  also  available  that  are  a 
combination  of  radiators  and  convectors. 

HEAT  EMISSION  OF  RADIATORS  AND  CONVECTORS 

All  heating  units  emit  heat  by  radiation  and  conduction.  The  resultant 
heat  from  these  processes  depends  upon  whether  or  not  the  heating  unit  is 
exposed  or  enclosed  and  upon  the  contour  and  surface  characteristics  of 
the  material  in  the  units. 

An  exposed  radiator  emits  less  than  half  of  its  heat  by  radiation,  the 
amount  depending  upon  the  size  and  number  of  sections.  When  the 
radiator  is  enclosed  or  shielded,  radiation  is  further  reduced.  The  balance 
of  the  emission  is  by  conduction  to  the  air  in  contact  with  the  heating 
surface*  and  the  resulting  circulation  of  the  air  warms  by  convection, 

A  built-in  heating  unit  in  a  con  vector  emits  practically  all  of  its  heat  by 
conduction  to  the  air  surrounding  it  and  this  heated  air  is  in  turn  trans- 
mitted by  convection  to  the  rooms  or  spaces  to  be  warmed,  the  heat 
emitted  by  radiation  being  negligible.  The  small  amount  of  heat  trans- 
mitted by  radiation  to  the  inside  surface  of  the  enclosure  diminishes  as  the 
surface  temperature  of  the  enclosure  approaches  the  surface  temperature 
of  the  heating  unit. 

TYPES  OF  RADIATORS 

Present  day  radiators  may  be  classified  as  tubular,  wall,  or  window 
types,  and  are  generally  made  of  cast  iron.  Catalogs  showing  the  many 
designs  and  patterns  available  now  include  a  junior  size  which  is  more 
compact  than  the  standard  unit. 

Pipe  Coil  Radiators 

Pipe  coils  are  assemblies  of  standard  pipe  or  tubing  (1  in.  to  2  in.)  which 
are  used  as  radiators.  In  older  practice  these  coils  were  commonly  used 
in  factory  buildings,  but  now  wall  type  radiators  are  most  frequently  used 

491 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


for  this  service.  When  coils  are  used,  the  miter  type  assembly  is  to  be 
preferred  as  it  best  cares  for  expansion  in  the  pipe.  Cast  manifolds  or 
headers,  known  as  branch  tee-s,  are  available  for  this  construction. 

OUTPUT  OF  RADIATORS 

The  output  of  a  radiator  can  be  measured  only  by  the  heat  it  emits. 
The  old  standard  of  comparison  used  to  be  square  feet  of  actual  surface, 
but  since  the  advance  in  radiator  design  and  proportions,  the  surface 
area  alone  is  not  a  true  index  of  output.  (The  engineering  unit  of  output 
is  now  the  Mb  or  1000  Btu.)  However,  during  the  period  of  transition 
from  the  old  to  the  new,  radiators  may  be  referred  to  in  terms  of  equivalent 
square  feet.  For  steam  service  this  is  based  on  an  emission  of  240  Btu 
per  hour  per  square  foot. 

TABLE  1.    VARIATION  IN  DIMENSIONS  AND  CATALOG  RATINGS  OF 
10-SECTiON  TUBULAR  RADIATORS 


No  of  Tubes 

3 

4 

5 

6 

7 

Width  of  Radiator.,  Inches 

4.6-5.1 

6.0-7.0 

8.0-8.9 

9.1-10.4 

11.4-12.8 

Length  per  Section  Inches 

2.5 

2.5 

2.5 

2.5 

2.5-3.0 

HEIGHT  WITH  LEGS  —  INCHES 

HEAT  EMISSION—  EQUIVALENT  SQUARE  FEET 

13-14 
16-18 
20-21 
22-23 
25-26 
30-32 
36-38 

20 

25.0-32.5 
30.0-38.3 
36.7-45.0 
40.0-45.2 
50.0-53.5 
63.3-62.5 
70.0-75.4 

28.5 

15.0-17.5 
20.0-21.3 
20.0-26.7 
25.0-30.9 
30.0-36.7 

20.0-22.5 
25 
25.0-27.5 
33.3-35.0 
40.0-4:2.5 

25.0-31.2 
30.0-33.9 
32.5-39.8 
40.0-48.6 
50.0-56.5 

30 
35 
37.5-40.0 
50 
60 

Output  of  Tubular  Radiators 

Table  1  illustrates  the  difficulty  in  tabulating  tubular  radiator  outputs 
since  there  is  so  much  variation  between  the  products  of  the  different 
manufacturers.  Only  on  the  four-tube  and  six-tube  sizes  is  there  any 
practical  agreement  in  output  value.  The  heat  emission  values  appear  as 
square  feet  but  are  entirely  empirical,  being  based  on  the  heat  emission  of 
the  radiator  and  not  on  the  measured  surface. 

Output  of  Wall  Radiators 

An  average  value  of  300  Btu  per  actual  square  foot  of  surface  area  per 
hour  has  been  found  for  wall  radiators  one  section  high  placed  with  their 
bars  vertical.  Several  recent  tests1  show  that  this  value  will  be  reduced 
from  5  to  10  per  cent  if  the  radiator  is  placed  near  the  ceiling  with  the  bars 
horizontal  and  in  an  air  temperature  exceeding  70  F.  When  radiators 
are  placed  near  the  ceiling,  there  is  usually  so  noticeable  a  difference  in 
temperature  between  the  floor  level  and  the  ceiling  that  it  becomes  dif- 
ficult to  heat  the  living  zone  of  a  room  satisfactorily. 


University  of  Illinois,  Engineering  Experiment  Station  Bulletin  No.  223,  p.  30. 

492 


CHAPTER  30 — RADIATORS  AND  GRAVITY  CONVECTORS 


Output  of  Pipe  Coils 

The  heat  emission  of  pipe  coils  placed  vertically  on  a  wall  with  the 
pipes  horizontal  is  given  in  Table  2.  This  has  been  developed  from  avail- 
able data  and  does  not  represent  definite  results  of  tests.  For  such  coils 
the  heat  emission  varies  as  the  height  of  the  coil.  It  is  customary  to  use 
an  average  emission  of  100  Btu  per  linear  foot  of  IJ^-in.  pipe,  10  ft  high. 
The  heat  emission  of  each  pipe  of  ceiling  coils,  placed  horizontally,  is  about 
126  Btu,  156  Btu,  and  175  Btu  per  linear  foot  of  pipe,  respectively,  for 
1-in.,  lj^-in.,  and  IJ^-in.  coils. 


TABLE  2.    HEAT  EMISSION  OF  PIPE  COILS  PLACED  VERTICALLY  ON  A  WALL  (PIPES 
HORIZONTAL)  CONTAINING  STEAM  AT  215  F  AND  SURROUNDED  WITH  AIR  AT  70  F 

Btu  per  linear  foot  of  coil  per  hour  (not  linear  feet  of  pipe) 


SIZE  OF  PIPE 

1  IN. 

IK  IN. 

IJilN 

Single  row 

132 

162 

185 

Two               .—             

252 

312 

348 

Four 

440 

545 

616 

Six.      

567 

702 

793 

Eight  

651 

796 

907 

Ten 

732- 

907 

1020 

Twelve  

812 

1005 

1135 

Effect  of  Paint 

The  prime  coat  of  paint  on  a  radiator  has  little  effect  on  the  heat  output, 
but  the  finishing  coat  of  paint  does  influence  the  radiation  emission.  Since 
this  is  a  surface  effect,  there  is  no  noticeable  change  in  the  convection  loss. 
Thus,  the  larger  the  proportion  of  direct  radiating  surface,  the  greater 
will  be  the  effect  of  painting  on  the  radiation.  Available  tests  are  on  old- 
style  column  type  radiators  which  gave  results  shown  in  Table  3. 

TABLE  3.    EFFECT  OF  PAINTING  32-iN.  THREE  COLUMN,  SIX-SECTION 
CAST-IRON  RADIATOR* 


RADIATOR 

FINISH 

AREA 

QA    TTrn 

COEFFICIENT 
OF  HEAT  TRANS. 

RELATIVE 
HEATING  VALUE 

BTU 

PER  CENT 

1 

Bare  iron,  foundry  finish     

27 

1.77 

100.5 

2 

One  coat  of  aluminum  bronze  

27 

1.60 

90.8 

3 

Gray  paint  dipped...-  

27 

1.78 

101.1 

4 

One  coat  dull  black  Pecora  paint.— 

27 

1.76 

100.0 

aComparative  Tests  of  Radiator  Finishes,  by  W.  H.  Severns  (A.S.H.V.E.  TRANSACTIONS,  Vol.  33, 1927). 

HEATING  EFFECT 

For  several  years  the  heating  effect  of  radiators  has  been  considered  by 
engineers  in  order  to  use  it  for  the  rating  of  radiators  and  in  the  design  of 
heating  systems.  Heating  effect  is  the  useful  output  of  a  radiator,  in  the 
comfort  zone  of  a  room,  as  related  to  the  total  input  of  the  radiator2. 


*The'  Seating  Effect  of  Radiators,  by  Dr.  Charles  Brabble  (A.S.H.V.E,  TRANSACTIONS,  Vol.  33,  1927. 
p.  33). 

493 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

The  results  of  tests  conducted  at  the  University  of  Illinois  are  shown  in 
Figs.  1  and  2s.  For  the  four  types  of  radiators  shown,  the  following  con- 
clusions are  given: 


-22 

-22 
-2.7 

-2? 


"<?/£/  f?oom 


ss      J4ZJ&.  Test- f?--£:/Q /-Tv&e-F&ne/ F&J. 
••     55O  ib.jTe^f  ff-Zc,  (&csf  223)  Wa/J  ffbcf. 


fiejgh?  Abo*?  rtoor  /n 


FIG.  1.    ROOM  TEMPERATURE  GRADIENTS  AND  STEAM  CONDENSING  RATES  FOR  FOUR 
TYPES  OF  CAST-IRON  RADIATORS  WITH  A  COMMON  TEMPERATURE  AT  THE  60-lN.  LEVEL 

Note  that  the  steam  condensations  are  practically  the  same  for  all  four  radiators  when  the  same  air 
temperature  of  69  F  i?  maintained  at  the  60-in.  level. 


6-48  /ty.  7&sf  f?-£20t  3 -Tube  ftae/- 
Test-  /?-£-S^-f-  T 


-tQ/2,343t7fflJO 


FIG.  2.    ROOM  TEMPERATURE  GRADIENTS  AND  STEAM  CONDENSING  RATES  FOR  FOUR 
TYPES  OF  CAST-IRON  RADIATORS  WITH  A  COMMON  TEMPERATURE  AT  THE  30-lN.  LEVEL 

Note  that  the  steam  condensations  are  different  for  all  four  radiators  when  the  same  air  temperature  of 
88  F  is  maintained  at  the  SO-in.  level. 


1.  The  heating  effect  of  a  radiator  cannot  be  judged  solely  by  the  amount  of  steam 
condensed  within  the  radiator. 

2.  Smaller  fioor-to-ceiling  temperature  differentials  can  be  maintained  with  long,  low, 
thin,  direct  radiators,  than  is  possible  with  high,  direct  radiators. 


'Steam  Condensation  an  Inverse  Index  of  Heating  Effect,   by  A.  P.   Kratz  and  M.   K.  Fahneatock 
(A.S.H.V.E.  TRANSACTIONS,  Vol.  37,  1931). 

494 


CHAPTER  30 — RADIATORS  AND  GRAVITY  CONVECTORS 

3.  The  larger  portion  of  the  floor-to-ceiling   temperature  differential  in  a  room  of 
average  ceiling  height  heated  with  direct  radiators  occurs  between  the  floor  and  the 
breathing  level. 

4.  The  comfort  level  (approximately  2  ft-6  in.  above  floor)  is  below  the  breathing  line 
level  (approximately  5  ft-0  in.  above  floor),  and  temperatures  taken  at  the  breathing 
line  may  not  be  indicative  of  the  actual  heating  effect  of  a  radiator  in  the  room.    The 
comfort-indicating  temperature  should  be  taken  below  the  breathing  line  level. 

5.  High  column  radiators  placed  at  the  sides  of  window  openings  do  not  produce  as 
comfortable  heating   effects   as  long,   low,    direct  radiators  placed  beneath  window 
openings4. 

HEATING  UP  THE  RADIATOR 

The  maximum  condensation  occurs  in  a  heating  unit  when  the  steam 
is  first  turned  on.  Fig.  3  shows  a  typical  curve  for  the  condensation  rate 
in  pounds  per  hour  for  the  time  elapsing  after  steam  is  turned  into  a  cast- 
iron  radiator.  The  data  are  from  tests  on  old  style  column  type  radiators. 


fate  of  Condensation  per5q.FtperHr.(inlbJ 

o  'o  8  8  fe  §  §  ~g  fe  5§  | 

/ 

\ 

/ 

\ 

/ 

\ 

/ 

^ 

\ 

/ 

\ 

/ 

\ 

/ 

\ 

\ 

,       > 

/ 

.^ 

/ 

I/ 

**•    "0  10  20  30  40 

Time     elapsing   after  Siecim  is  turned  in+o  Radiortor  (m  Minutes! 

FIG.  3.   CHART  SHOWING  THE  STEAM  DEMAND  RATE  FOR  HEATING  UP  A  CAST-IRON 
RADIATOR  WITH  FREE  AIR  VENTING  AND  AMPLE  STEAM  SUPPLY 

In  practice  the  rate  of  steam  supply  to  the  heating  unit  while  heating  up 
is  frequently  retarded  by  controlled  elimination  of  air  through  air  valves 
or  traps.  Automatic  control  valves  may  also  retard  the  supply  of  steam. 

ENCLOSED  RADIATORS 

The  general  effect  of  an  enclosure  placed  about  a  direct  radiator  is  to 
restrict  the  air  flow,  diminish  the  radiation  and,  when  properly  designed, 
improve  the  heating  effect.  Recent  investigations5  indicate  that  in  the 
design  of  the  enclosure  three  things  should  be  considered : 

1.  There  should- be  better  distribution  of  the  heat  below  the  breathing  line  level  to 
produce  greater  heating  comfort  and  lowered  ceiling  temperatures. 


*Effect  of  Two  Types  of  Cast  Iron  Steam  Radiators  in  Room  Heating,  by  A.  C.  Willard  and  M.  K. 
Fahnestock  (Heating,  Piping,  and  Air  Conditioning,  March,  1930). 

•University  of  Illinois  Engineering  Experiment  Station  Bulletins  No.  192  and  223,  and  Investigation  of 
Heating  Rooms  with  Direct  Steam  Radiators  Equipped  with  Enclosures  and  Shields,  by  A.  C.  Willard, 
A.  P.  Kratz,  M.  K.  Fahnestock  and  S.  Konzo  (A.S.H.V.E.  TRANSACTIONS,  Vol.  35,  1929), 

495 


AMERICAN  SOCIETY  of  HEATING  4nd  VENTILATING  ENGINEERS  GUIDE,  1935 

2.  The  lessened  steam  consumption  may  not  materially  change  the  radiator  heating 
performance. 

3.  The  enclosed  radiator  may  inadequately  heat  the  space. 

A  comparison  between  a  bare  or  exposed  radiator  (A)  and  the  same 
radiator  with  a  well-designed  enclosure  (j5)r  with  a  poorly-designed 
enclosure  (C),  and  with  a  cloth  cover  (D)  will  illustrate  the  relative 
heating  effects.  In  Fig.  4  the  curve  (B)  reveals  that  the  enclosed  radiator 
used  less  steam  than  the  exposed  radiator,  but  gave  a  satisfactory  heating 
performance.  A  well-designed  shield  placed  over  a  radiator  gives  about 
the  same  heating  effect.  Curve  (C)  shows  the  unsatisfactory  effects 
produced  by  improperly  designed  enclosures.  Curve  (D)  shows  that  the 
effect  of  a  cloth  cover  extending  downward  6  in.  from  the  top  of  the 
radiator  was  to  make  the  performance  unsatisfactory  and  inadequate. 


TEMPERATURE  IN  DEG  FAHR 
8  8  3  8 

A- 

0< 

^ 

,  

^—  •* 

-^* 

^w-+ 

B 

IS 

v< 

&*A 

fc^- 

^-"* 

^ 

£ 

if 

X 

f^ 

c 

<& 

'/& 
'  *S 

?X 

.x" 

Radiator 

Steam  consumption 

Lb  per  hr 

Per  cent 

*    , 
/ 

/ 

A 
B 
C 
D 

5.44 
4.71 
4.50 
4.59 

100 
86.6 
82.7 
84.4 

— 

)                2                4                6                8               1 

HEIGHT  ABOVE  FLOOR  IN  FEET 

FIG.  4.  STEAM  CONSUMPTION  OF  EXPOSED  AND  CONCEALED  RADIATORS 

Practically  all  commercial  enclosures  and  shields  for  use  on  direct 
radiators  are  equipped  with  water  pans  for  the  purpose  of  adding  moisture 
to  the  air  in  the  room.  Tests6  show  that  an  average  evaporative  rate  of 
about  0.235  Ib  per  square  foot  of  water  surface  per  hour  may  be  obtained 
from  such  pans,  when  the  radiator  is  steam  hot  and  the  relative  humidity 
in  the  room  is  between  25  and  40  per  cent.  This  source  of  supply  of 
moisture  alone  is  not  adequate  to  maintain  a  relative  humidity  above 
25  per  cent  on  a  zero  day. 

CONVECTORS  OR  CONCEALED  HEATERS 

Although  any  standard  heating  unit  (i.e.,  radiator)  may  be  concealed 
in  a  cabinet  or  other  enclosure  so  that  the  greatest  percentage  of  heat  is 


"University  of  Illinois  Engineering  Experiment  Station  Bulletin  No.  230,  p,  20. 

496 


CHAPTER  30 — RADIATORS  AND  GRAVITY  CONVECTORS 

conveyed  to  the  room  by  convection,  the  best  results  are  usually 
obtained  where  units  of  special  design  are  used.  Commercially,  these 
specially  designed  units  are  built  in  as  part  of  the  enclosing  cabinets  which 
are  necessary  for  the  proper  functioning  of  these  heaters.  As  distin- 
guished from  radiators,  these  gravity  con  vectors  have  come  to  be  known  as 
concealed  heaters.  Fig.  5  shows  a  typical  built-in  cabinet  convector. 


SECTION 

FIG.  5.    TYPICAL  CONCEALED  CONVECTOR  USING  SPECIALLY  DESIGNED  HEATING  UNIT 

'  The  elements  or  heating  units  usually  consist  of  a  relatively  large  amount 
of  extended  surface  which  may  be  integral  with  the  core  or  assembled  over 
it,  making  thermal  contact  by  pressure,  through  solder,  or  by  both  pres- 
sure and  metallic  contact.  Heating  elements  may  be  of  cast-iron,  cast 
aluminum,  sheet  steel,  copper,  or  commercial  alloys. 

Concealed  heaters  or  convectors  maintain  room  temperatures  with  low 
steam  consumption  due,  probably,  to  their  performance  characteristics 
which  give  reduced  air  temperatures  in  the  upper  level  of  a  room  with  a 
directed  flow  of  warm  air  into  the  living  zone  and  but  little  radiant  heat  to 
exposed  surfaces.  The  Concealed  Heater  Manufacturers  Association  has 
decided  to  use  the  A.S.H.V.E.  Standard7  in  the  formulation  of  its  ratings, 
but  has  made  a  provision  that  heating  effect  be  included  in  the  ratings  in 
accordance  with  the  following  rules: 


"  'A.S.H.V.E.  Standard  Code  for  Testing  and  Rating  Concealed  Gravity  Type ,  Radiation  (A.S.H.V.E. 
TRANSACTIONS,  Vol.  37,  1931). 

497 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

Con  vectors  20  in.  or  lower  may  have  a  heating  effect  of  15  per  cent  included  in  their 
rating;  for  higher  convectors,  1  per  cent  shall  be  deducted  for  each  1  in.  increase  in 
height  above  20  in.  No  heating  effect  shall  be  included  in  the  ratings  for  convectors 
with  a  cabinet  35  in.  from  the  floor  to  the  top,  or  higher.  All  ratings  published  will  show 
definitely  that  this  heating  effect  factor  is  included  in  the  catalog  ratings. 

Concealed  heaters  or  convectors  are  generally  sold  as  completely 
built-in  units.  The  enclosing  cabinet  should  be  designed  with  suitable 
air  inlet  and  outlet  grilles  to  give  the  heating  element  its  best  performance. 
Tables  of  capacities  are  catalogued  for  various  lengths,  depths  and  heights, 
and  combinations  are  available  in  several  styles  for  installations,  such  as 
the  wall  hanging  type,  free-standing  floor  type,  recess  type  set  flush 
with  wall  or  offset,  and  the  completely  concealed  type.  Most  of  these 
types  may  be  arranged  with  a  top  outlet  grille,  although  the  front  outlet 
is  practically  standard.  In  cases  where  enclosures  are  to  be  used  but  are 
not  furnished  by  the  heater  manufacturer,  it  is  important  that  the  pro- 
portions of  the  cabinet  and  the  grilles  be  so  designed  that  they  will  not 
impair  the  performance  of  the  assembled  convector. 

The  output  of  a  concealed  heater,  for  any  given  length  and  depth,  is  a 
variable  of  the  height.  Published  ratings  are  generally  given  in  terms  of 
equivalent  square  feet,  corrected  for  heating  effect.  However,  an  extended 
surface  heating  unit  is  entirely  different  structurally  and  physically  from 
a  direct  radiator  and,  since  it  has  no  area  measurement  corresponding  to 
the  heating  surface  of  a  radiator,  many  engineers  believe  that  the  per- 
formance of  convectors  should  be  stated  in  Btu's.  For  steam  convectors, 
as  for  radiators,  240  Btu  per  hour  may  be  taken  as  an  equivalent  square 
foot  of  radiation. 

CODE  TESTS  FOR  RADIATORS  AND  CONVECTORS 

As  previously  indicated,  the  output  of  radiators  and  convectors  is  still 
designated  by  the  terms  of  older  practice,  but  this  is  gradually  giving  place 
to  an  engineering  method  of  designating  heat  emission.  The  A.S.H.V.E. 
has  adopted  the  following  standards :  Code  for  Testing  Radiators  (1927) ; 
Codes  for  Testing  and  Rating  Concealed  Gravity  Type  Radiation  (Steam, 
1932,  and  Hot  Water,  1933). 

For  steam  services  the  actual  condensation  weight  is  taken  without  any 
allowance  for  heating  effect;  for  hot  water  services  the  weight  of  circulated 
water  is  used  without  allowance  for  heating  effect.  In  all  cases  the  total 
heat  transmission  varies  as  the  1.3  power  of  the  temperature  difference 
between  that  inside  the  radiator  and  the  air  in  the  room,  and  is  expressed 
in  Btu  or  Mb  per  hour. 

Standard  test  conditions  specify  either  a  steam  pressure  of  1  Ib  gage 
(215  F),  or  hot  water  at  170  F  and  a  room  temperature  of  70  F  for  radi- 
ators, or  an  inlet  air  temperature  of  65  F  for  convectors.  The  heating 
capacity  of  a  steam  radiator  or  steam  convector  is  determined  as  follows: 

Ht  -  Wshfg  (1) 

where 

Ht  =  Btu  per  hour  under  test  conditions. 
W$  =s  condensation  in  Ib  per  hour. 
hfg  —  latent  heat  in  Btu  per  Ib. 

498 


CHAPTER  30 — RADIATORS  AND  GRAVITY  GONVECTORS 

Ht  may  be  converted  to  standard  conditions  of  code  ratings  by  using 
the  proper  correction  factor  from  the  following  formulae : 

For  radiators  : 

c  =  P15  -  7oy-3  =  /    145     y-8 

For  con  vectors: 

r    _    /215-65\1.3  _  /      150       \1.3 

t-s   —     I  ~~m          ^r~  }         —   I  ~^ ^r~  I  (*>) 

\  TS  -  TI  )        \  rs  -  TI  ) 

The  output  under  standard  conditions  will  be : 

Hs  —  Cs  Ht  ^4} 

"where 

Cs  =  correction  factor. 

Ts  =  steam  temperature  during  test,  degrees  Fahrenheit. 

IV  =  room  temperature  during  test,  degrees  Fahrenheit. 

T{  =  inlet  air  temperature  during  test,  degrees  Fahrenheit. 

H3  —  heat  emission  rating  under  standard  conditions,  Btu  per  hour. 

Similarly,  for  hot  water  convenors,  the  output  under  test  conditions  may 
be  determined  as  follows: 


H=W  (6!  -  02)  ^~  (5) 

where 

H  =  Btu  per  hour  under  test  conditions. 
W  —  pounds  of  water  handled  during  test. 
6j  —  average  temperature  of  inlet  water,  degrees  Fahrenheit. 
62  ss  average  temperature  of  outlet  water,  degrees  Fahrenheit. 
/  =  duration  of  test,  seconds. 

To  convert  test  results  to  standard  conditions,  the  following  correction 
factor  is  used: 

170-65       \i-3         /  105 


6!  -  6,  }  (6) 

— 


It  has  been  shown  that  when  the  exponent  1.3  is  used  the  range  of  error 
Is  less  than  5  per  cent8. 

GRAVITY-INDIRECT  HEATING  SYSTEMS9 

The  heating  units  for  this  system  are  usually  of  the  extended  surface 
type  for  steam  or  hot  water,  and  are  installed  about  as  shown  in  Fig.  6. 


treats  of  Convectors  in  a  Warm  Wall  Testing  Booth,  by  A.  P.  Kratz,  M.  K.  Fahnestock,  and  E.  L. 
Broderick  (Heating*  Piping  and  Air  Conditioning,  August,  1933) . 

>For  further  information  on  this  subject  see  A.S.H.V.E.  Code  of  Minimum  Requirements  for  the  Heating 
and  Ventilation  of  Buildings  (edition  of  1929)  and  Mechanical  Equipment  of  Buildings*  by  Harding  and 
Wfflard,  Vol.  I,  second  edition,  1929. 

499 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


Supports  hung  from  joist  o'r  floor  above 


FIG.  6.    GRAVITY-INDIRECT  HEATING  SYSTEM* 

aSee  Mechanical  Equipment  of  Buildings,  by  Harding  and  Willard,  Vol.  I,  second  edition,  1929. 

The  temperature  and  volume  of  the  air  leaving  the  register  must  be  great 
enough  so  that  in  cooling  to  room  temperature  the  heat  available  will  just 
equal  the  heat  loss  during  the  same  time.  In  cases  where  ventilation  is  a 
requirement,  the  air  volume  needed  may  become  so  large  that  the  entering 
air  temperature  will  be  but  slightly  above  the  room  temperature.  To 
establish  and  maintain  a  constant  heat  flow,  provision  must  be  made  for 
removing  the  air  in  the  room,  after  it  has  cooled  to  the  desired  room  tem- 
perature, by  a  system  of  vent  flues  or  ducts.  As  the  air  flow  is  maintained 
by  natural  draft  and  this  gravity  head  is  very  slight,  it  is  necessary  to 
make  all  ducts  as  short  as  possible,  especially  the  runs  from  the  heating 
units  to  the  base  of  the  vertical  warm  air  flues.  Gravity-indirect  arrange- 
ments, such  as  illustrated  in  Fig.  6,  are  not  to  be  generally  recommended 
for  hot  water  systems  unless  the  water  temperature  can  be  maintained  at 
a  reasonably  high  temperature  and  rapid  circulation  of  the  water  can  be 
had. 

PROBLEMS  IN  PRACTICE 

1  •  What  are  the  principal  differences  between  a  radiator  and  a  convector? 

A  radiator  Is  commonly  thought  of  as  a  commercial  heating  unit  having  a  maximum 
amount  of  direct  heating"  surface,  whereas  a  convector  is  a  heating  device  in  which  the 
extended  or  secondary  surface  may  be  several  times  that  of  the  prime  surface.  The 
radiator  ordinarily  has  vertical  tubular  chambers  for  the  heating  medium  but  most 
convectors  have  horizontal  tubular  chambers  to  which  fins  are  attached  so  as  to  form 

500 


CHAPTER  30 — RADIATORS  AND  GRAVITY  CONVECTORS 

vertical  flues  for  the  passage  of  air.  While  radiators  are  either  exposed,  enclosed,  or 
shielded,  convectors  are  concealed  by  means  of  a  tight-fitting  enclosure.  Radiators  are 
commonly  made  of  cast-iron  but  convectors  may  be  made  of  a  combination  of  metals, 
such  as  copper  and  brass,  or  copper  and  aluminum. 

2  •  How  did  the  term  heating  effect  come  into  use? 

It  has  been  found  that  a  room  requiring  a  radiator  of  a  certain  determined  capacity 
could  under  certain  conditions  be  properly  heated,  with  less  temperature  gradient  be- 
tween floor  and  ceiling  and  with  less  steam  condensation,  by  the  same  radiator  or  by  one 
of  a  different  design  having  the  same  commercially  rated  capacity.  This  resulted  in  the 
use  of  the  term  heating  effect  to  apply  to  the  useful  heat  output  of  a  radiator,  in  the  com- 
fort zone  of  a  room,  as  related  to  the  total  input  to  the  radiator. 

3  •  What  is  the  effect  of  enclosing  a  direct  radiator? 

This  will  depend  almost  entirely  on  the  design  of  the  enclosure.  If  properly  enclosed,  a 
radiator  can  be  made  to  give  better  heat  distribution  below  the  breathing  line  and  to 
condense  less  steam  than  does  an  unenclosed  radiator  giving  equal  comfort. 

4  •  Does  paint  on  a  direct  radiator  affect  its  heat  output? 

Aluminum  or  gold  bronze  paint  tends  to  reduce  the  heat  output  of  a  direct  radiator 
perhaps  10  per  cent,  but  ordinary  non-metallic  paint  will  have  little  effect  on  the  heat 
output. 

5  •  How  can  the  temperature  gradient  between  the  floor  and  the  ceiling  of  a 
Toom  he  maintained? 

Long,  low,  thin,  direct  radiators  will  maintain  smaller  floor  to  ceiling  temperature 
differences  than  high  direct  radiators.  Convectors  properly  selected  and  properly 
installed  will  accomplish  the  same  result. 

6  •  Is  the  method  of  enclosing  a  direct  radiator  different  from  that  required 
for  a  convector? 

Generally,  yes.  An  enclosure  for  a  direct  radiator  should  provide  a  space  of  at  least 
2  in.  between  the  radiator  and  the  front  and  back  inside  vertical  surfaces  of  the  enclosure 
to  utilize  the  radiant  heat  to  best  advantage  in  heating  the  air  stream  passing  through 
the  enclosure.  The  enclosure  for  a  convector  should  be  constructed  with  as  little  clearance 
as  possible  between  the  inside  vertical  surfaces  and  the  convector  so  as  to  confine  the 
passage  of  the  air  stream  through  the  fins  of  the  convector.  An  all-over  face  grille,  often 
used  when  direct  radiators  are  concealed,  should  never  be  used  for  a  convector.  The 
•essential  requirements  for  a  convector  enclosure  are  an  air  inlet  below  the  convector,  a 
warm  air  outlet  above  it,  and  sufficient  height  between  the  openings  to  provide  a  stack 
•effect. 

7  •  Is  it  necessary  to  make  any  allowance  for  the  performance  of  a  convector 
hecause  it  is  enclosed? 

No.  The  commercial  ratings  of  convectors  have  been  determined  by  testing  the  con- 
vectors  in  proper  enclosures  with  grilles  in  place  just  as  they  should  be  installed  for 
ordinary  service. 

$  •  On  what  hasis  are  the  capacities  of  convectors  published? 

Published  ratings  of  convectors  are  on  the  basis  of  equivalent  square  feet  of  direct  exposed 
cast-iron.  If  any  allowance  is  made  for  heating  effect,  the  amount  of  such  allowance  is 
generally  stated  in  the  manufacturers'  catalogs. 

9  •  How  are  fins  of  convectors  attached  to  the  tubes  or  prime  surface? 

Tubes  or  a  solid  core  may  be  forced  through  piercings  in  the  fins  under  pressure,  or  the 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

tubes  may  be  expanded  into  the  holes  through  the  fins.  In  addition  a  metallic  bonding 
agent  is  sometimes  used  to  insure  permanent  contact. 

10  •  What  is  the  procedure  in  selecting  a  convector  when  the  required  amount 
of  radiation  is  known? 

First  the  limiting  factor  or  factors  of  the  enclosure  must  be  determined  so  the  available 
size  of  the  wall  recess  can  be  found.  Manufacturers'  catalogs  show  capacities  of  con- 
vectors  of  each  standard  length  and  depth  with  varying  enclosure  heights.  From  these 
capacity  tables,  the  proper  convector  of  the  required  capacity  can  be  selected  for  the 
available  wall  recess.  If  all  three  dimensions  of  the  wall  recess  are  insufficient  to  accom- 
modate a  convector  of  the  required  capacity,  the  available  height  and  length  can  be 
maintained,  but  greater  depth  can  be  obtained  by  using  a  partially  recessed  enclosure. 

11  •  Given  a  room  to  be  heated  to  80  F  with  outside  temperature  at  zero  F. 
Assume  the  heat  loss  under  these  conditions  to  be  10,000  Btu  per  hour.     Deter- 
mine the  size  of  the  steam  radiator  to  be  installed. 

A  square  foot  of  radiation  is  equivalent  to  a  heat  emission  of  240  Btu  per  hour  under 
standard  conditions  of  steam  at  one  pound  gage  pressure  (215  F)  and  surrounding  air 
at  70  F.  With  surrounding  air  at  80  F,  the  heat  emission  from  a  radiator  will  be  less. 
Under  these  conditions,  the  heat  emission  will  not  be  240  Btu  per  square  foot  of  catalog 
rating  per  hour,  but  240  Cs. 

_    («.  -  'i)1'3 
~~ 


-  70)     ~     ' 

and  240  Cs   =  240  X  0.912   =  218.5  Btu.     Therefore,  the  size  of  the  radiator  to  be 
selected  shall  have  a  catalog  rating  of  10,000  divided  by  218.5  or  45.8  sq  ft. 


Chapter  31 

STEAM  HEATING  SYSTEMS 

Gravity  and  Mechanical  Return,  Gravity  One-Pipe  Air-Vent 
System,  Gravity  Two-Pipe  Air 'Vent  System,  One-Pipe  Vapor 
System,  Two-Pipe  Vapor  System,  Atmospheric  System,  Vacuum 
System,  Sub -Atmospheric  System,  Orifice  System,  Zone  Control, 
Condensation  Return  Pumps,  Vacuum  Pumps,  Traps 

THE  essential  features  of  the  common  type  of  steam  heating  systems 
are  described  in  this  chapter.  They  may  be  classified  according  to  the 
piping  arrangement,  the  accessories  used,  the  method  of  returning  the  con- 
densate to  the  boiler,  the  method  of  expelling  air  from  the  system,  or  the 
type  of  control  employed.  Information  concerning  the  design  and  layout 
of  steam  heating  systems  will  be  found  in  Chapter  32. 

GRAVITY  AND  MECHANICAL  RETURN 

In  gravity  systems  the  condensate  is  returned  to  the  boiler  by  gravity 
due  to  the  static  head  of  water  in  the  return  mains.  The  elevation  of  the 
boiler  water  line  must  consequently  be  sufficiently  below  the  lowest 
heating  units  and  steam  main  and  dry  return  mains  to  permit  the  return 
of  condensate  by  gravity.  The  water  line  difference1  must  be  sufficient  to 
overcome  the  maximum  pressure  drop  in  the  system  and,  when  radiator 
and  drip  traps  are  used  as  in  two-pipe  vapor  systems,  the  operating 
pressure  of  the  boiler.  This  applies  only  to  closed  circuit  systems,  where 
the  condensation  is  returned  to  the  boiler.  If  the  condensation  is  wasted, 
no  water  line  difference  is  required. 

In  mechanical  systems  the  condensate  flows  to  a  receiver  and  is  then 
forced  into  the  boiler  against  the  boiler  pressure.  The  lowest  parts  of  the 
supply  side  of  the  system  must  be  kept  sufficiently  above  the  water  line 
of  the  receiver  to  insure  adequate  drainage  of  water  from  the  system,  but 
the  relative  elevation  of  the  boiler  water  line  is  unimportant  in  such  cases 
except  that  the  head  on  the  pump  or  trap  discharge  becomes  greater  as 
the  height  of  the  boiler  water  line  above  the  trap  or  pump  increases. 

There  are  three  general  types  of  mechanical  returns  in  common  use, 
namely,  (1)  the  mechanical  return  trap,  (2)  the  condensation  return 
pump,  and  (3)  the  vacuum  return  pump.  Further  information  on  pumps 
and  traps  will  be  presented  later  in  this  chapter, 

GRAVITY  ONE-PIPE  AIR-VENT  SYSTEM 

In  the  gravity  one-pipe  air-vent  system  each  radiator  has  but  a  single 
connection  through  which  steam  must  enter  and  condensation  must 


The  water  line  difference  is  the  distance  between  the  water  line  of  the  boiler  and  the  low  point  of  the 
water  in  the  dry  return  main. 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

return  in  the  opposite  direction.     Each  radiator  has  an  individual  air 
valve. 

Up- Feed  Gravity  One- Pipe  Air- Vent  System 

This  system  is  the  most  common  of  all  methods  of  steam  heating,  due 
largely  to  its  low  cost  of  installation  and  its  simplicity.  As  will  be  seen 
from  Fig.  1,  the  steam  piping  rises  to  a  point  as  high  as  possible  at  the 
boiler  and  pitches  downward  from  this  location  until  the  far  end  of  the 
main  or  mains  is  reached.  At  the  far  ends  drips  are  taken  off  at  the  low 
points  of  the  steam  mains,  are  water-sealed  below  the  boiler  water  line, 
and  then  brought  back  to  the  boiler  in  a  wet  return.  Single  pipe  risers 


AIR  VALVE 


HARTFORD 
RETURM 
COMMECTlQh- 


FIG.  1.    TYPICAL  UP-FEED  GRAVITY  ONE-PIPE  AIR- VENT  SYSTEM 


are  branched  off  the  main  or  mains  to  feed  the  radiators,  the  steam  passing 
up  the  riser  and  the  condensation  flowing  down  it.  The  steam  and  con- 
densation flow  in  opposite  directions  in  the  riser  but  after  the  condensa- 
tion enters  the  steam  main  it  flows  in  the  same  direction  as  the  steam  and 
is  disposed  of  through  the  drip  connection  at  the  end  of  the  main.  In 
buildings  of  several  stories,  it  is  customary  to  drip  the  heel  of  each  riser 
separately,  whereas  in  one-  or  two-story  buildings  this  is  not  necessary. 
Both  types  of  branches  and  risers  are  shown  in  Fig.  1. 

Horizontal  branches  to  radiators  and  risers  should  be  pitched  at  least 
Yz  in.  in  10  ft  downward  toward  the  riser  or  vertical  pipe,  and  the  hori- 
zontal branches  from  the  steam  main  should  be  graded  at  least  this 
amount  toward  the  main,  except  where  the  heel  of  the  riser  is  dripped,  in 
which  case  the  branch  should  pitch  down  toward  the  riser  drip  (Figs.  2 
and  3).  The  return  line,  if  wet,  may  be  run  without  pitch  or  may  be 
pitched  in  either  direction,  but  if  it  is  necessary  to  carry  the  return  main 

504 


CHAPTER  31 — STEAM  HEATING  SYSTEMS 


overhead  for  any  distance  before  dropping,  the  return  should  slope  down- 
ward with  the  flow. 

The  radiator  valves  may  be  of  the  angle-globe  or  gate  type.  They 
should  not  be  of  the  straight-globe  type  because  the  damming  effect  of  the 
raised  valve  seat  interferes  with  the  flow  of  condensation  through  the 
valve.  Graduated  valves  cannot  be  used,  as  the  steam  valves  on  this 
system  must  be  fully  open  or  closed  to  prevent  the  radiators'  filling  with 
water.  Air  valves  may  be  manual  or  automatic,  with  or  without  a  check 
to  prevent  the  re-entrance  of  expelled  air.  Usually  the  automatic  type  is 
installed.  The  greatest  source  of  difficulty  with  one-pipe  steam  systems 
is  that  the  heat  is  all  on  or  all  off,  with  no  intermediate  position  possible. 
However,  intelligent  use  of  the  on-and-off  method  of  manual  control 
gives  reasonably  satisfactory  results. 

It  is  important  that  the  lowest  points  of  the  steam  mains  and  heating 
units  be  kept  sufficiently  above  the  water  line  of  the  boiler  to  prevent 


FIG.  2.     TYPICAL  STEAM  RUNOUT  WHERE      FIG.  3.    TYPICAL  STEAM  RUNOUT  WHERE 
RISERS  ARE  NOT  DRIPPED  RISERS  ARE  DRIPPED 

flooding,  although  proper  design  will  eliminate  this  danger.  Usually  18 
in.  is  sufficient  but  construction  limitations  frequently  make  shorter  dis- 
tances necessary.  The  distance  may  be  checked  in  the  following  manner : 

Referring  to  Fig.  4  it  will  be  seen  that  the  water  in  the  wet  return  is  really  in  an  in- 
verted siphon,  or  U-shaped  container,  with  the  boiler  steam  pressure  on  the  top  of  the 
water  at  one  end  and  the  steam  main  pressure  on  the  top  of  the  water  at  the  other  end. 
The  difference  between  these  two  pressures  is  the  pressure  drop  in  the  system,  i.e.,  the 
friction  of  the  steam  in  passing  from  the  boiler  to  the  far  end  of  the  main.  The  water  in 
the  far  end  will  rise  sufficiently  to  overcome  this  difference  in  order  to  balance  the  pres- 
sures, and  it  will  rise  enough  farther  to  produce  a  flow  through  the  return  into  the  boiler 
(usually  about  3  in.  unless  the  pipes  are  small  or  full  of  sediment),  and  it  will  rise  still 
farther  if  a  check  valve  is  installed  in  the  return  so  as  to  obtain  sufficient  head  to  lift  the 
tongue  of  the  check  (usually  4  in.  will  be  necessary). 

If  a  one-pipe  steam  system  is  designed,  for  example,  for  a  total  pressure  drop  of  J^  Ib, 
and  utilizes  an  Underwriters  Loop2  instead  of  a  check  valve  on  the  return,  the  rise  in  the 
water  level  at  the  far  end  of  the  return  due  to  the  difference  in  steam  pressure  would  be 
Y%  of  28  in.,  or  3  M  in.  Adding  3  in.  to  this  for  the  flow  through  the  return  main  and  6  in. 
as  a  factor  of  safety  gives  12 y%  in.  as  the  distance  the  bottom  of  the  lowest  part  of  the 
steam  main  and  all  heating  units  must  be  above  the  boiler  water  line.  The  same  system, 
however,  installed  and  sized  for  a  total  pressure  drop  of  J^  Ib,  and  with  a  check  in  the 
return,  would  require  J^  of  28  in.,  or  14  in.,  for  the  difference  in  steam  pressure,  3  in.  for 
the  flow  through  the  return,  4  in.  to  operate  the  check,  and  6  in.  for  a  factor  of  safety, 
making  a  total  of  27  in.  as  the  required  distance.  Higher  pressure  drops  would  increase 
the  distance  accordingly. 


5See  discussion  of  piping  details  in  Chapter  32. 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


-Steam  main 


Boiler  steam  pressure 


t-rot 


Steam  pressure  at 
end  of  main 

Return  water 


FIG.  4.    DIFFERENCE  IN  STEAM  PRESSURE 

ON  WATER  IN  BOILER  AND  AT  END 

OF  STEAM  MAIN 


Steam  drop  to  radiators  - 

FIG.  6.   STEAM  RUNOUTS  DRIPPING  MAIN 

Pitch—*  _  Runout  x 


-Steam  mam 


stearn  drop  to  rad^o 


FIG.  7.    STEAM  RUNOUTS  WITH  MAIN 
DRIPPED  AT  END  ONLY 


Down-Feed  Gravity  One-Pipe  Air-Vent  System 

In  the  overhead  down-feed  gravity  one-pipe  air- vent  system  there  is  no 
change  over  the  up-feed  system  in  the  radiators,  the  radiator  valves,  the 
air  valves,  or  the  radiator  runouts  as  far  back  as  the  risers.  Beyond  this 
point  there  are  basic  differences'.  The  steam  is  taken  from  the  boiler  and 
carried  to  the  top  of  the  building  as  near  the  boiler  as  possible  (Fig.  5). 
If  the  run  to  the  main  riser  is  long,  or  if  the  riser  extends  several  stories  in 
order  to  reach  the  top,  the  bottom  of  the  riser  should  be  dripped  into  the 
wet  return.  The  horizontal  main  is  taken  off  the  top  of  the  riser  and 
grades  down  from  the  riser  toward  all  of  the  drops,  each  drop  taking  its 
share  of  the  main  condensation  (Fig.  6),  or  all  of  the  drops  except  the  last 
may  be  taken  from  the  top  of  the  main  (Fig.  7) ,  the  last  drop  being  from 
the  bottom  and  serving  as  a  drain  for  the  entire  main.  As  the  overhead 


SUPPLY  RISEJ^ 


HARTFORD 

RETURN 
CONNECTION 


FIG.  5.    TYPICAL  DOWN-FEED  GRAVITY' ONE- PIPE  AIR- VENT  SYSTEM 

506 


CHAPTER  31 — STEAM  HEATING  SYSTEMS 


main  does  not  carry  any  condensation  from  the  radiators  it  is  immaterial 
which  method  is  used.  The  air  vent  shown  on  the  main  just  before  the 
last  drop  (Fig.  5)  may  be  placed  at  this  point  or  it  may  be  located  at  the 
bottom  of  the  drop  under  the  last  radiator  connection  and  sufficiently 
above  the  water  line  of  the  boiler  to  prevent  flooding. 

GRAVITY  TWO-PIPE  AIR-VENT  SYSTEM 

The  gravity  two-pipe  system  is  now  considered  obsolete  although  many 
of  these  systems  are  still  in  use  in  older  buildings.    Separate  supply  and 


AIR  VALVE 


HARTFORD 
RETURM 
COHNECTIi 


FIG.  8.    TYPICAL  UP-FEED  GRAVITY  TWO-PIPE  AIR- VENT  SYSTEM 


return  mains  and  connections  are  required  for  each  heating  unit;  air 
valves  are  installed  on  the  heating  units  and  mains;  hand  valves  are 
installed  on  the  returns. 

Up-Feed  Gravity  Two-Pipe  System 

This  system  (Fig.  8)  has  a  steam  and  a  return  connection  to  each 
radiator.  The  radiator  valves  for  steam,  return,  and  air  are  the  same  as 
those  described  for  the  gravity  one-pipe  air-vent  system.  The  steam 
main  is  run  and  pitched  in  the  same  manner  as  in  the  one-pipe  system, 
but  the  returns  from  each  radiator  are  connected  into  a  separate  return 
line  system  which  has  its  risers  carried  down  and  joined  to  a  wet  return 
line  under  the  boiler  water  line  level.  Where  the  return  has  to  be  kept 
high  to  function  as  a  dry  return,  it  is  advisable  to  connect  the  return 
risers  to  the  dry  return  main  through  water  seals  about  36  in.  deep,  as 
shown  in  Fig.  9,  to  prevent  steam  from  one  riser  entering  another  and 
closing  the  air  valves  on  the  nearest  radiators. 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

Down -Feed  Gravity  Two- Pipe  System 

The  steam  main  in  the  down-feed  system  is  carried  to  the  top  of  the 
building,  and  the  piping  of  the  steam  side  is  arranged  practically  as  in  the 
down-feed  one-pipe  gravity  system.  The  drips  at  the  bottoms  of  the 
steam  drops  and  the  runouts  to  the  radiators  are  similar  to  those  shown 
in  Fig.  8  for  the  up-feed  gravity  two-pipe  system.  On  the  return  side  of 
the  system,  the  piping  is  arranged  in  exactly  the  same  manner  as  the 
up-feed  gravity  two-pipe  system. 

ONE-PIPE  VAPOR  SYSTEM 

A  vapor  system  is  one  which  operates  under  pressures  at  or  near 
atmospheric  and  which  returns  the  condensation  to  the  boiler  by  gravity. 
The  piping  arrangement  of  a  one-pipe  vapor  system  is  similar  to  that  of 


Return  from  radiators-*^ 
-Pitch 


Clean  out  ^J± 


FIG.  9.    METHOD  OF  CONNECTING  TWO-PIPE  GRAVITY  RETURNS  TO 
DRY  RETURN  MAIN 


the  gravity  one-pipe  steam  system ;  in  fact,  one-pipe  gravity  installations 
may  readily  be  changed  to  one-pipe  vapor  systems  by  making  a  few 
simple  alterations.  The  steam  radiator  valve  is  a  plug  cock  which  when 
opened  gives  a  free  and  unobstructed  passageway  for  water.  The  auto- 
matic air  valve  is  of  special  design  to  permit  the  ready  release  of  air  from 
the  radiator  and  to  prevent  the  return  of  the  air  after  it  is  expelled.  The 
air  valves  on  the  main  are  a  quick  relief  type,  and  the  whole  system  is 
designed  to  operate  on  a  few  ounces  of  pressure. 

TWO-PIPE  VAPOR  SYSTEM 

Two-pipe  vapor  systems  may  be  classified  as  (1)  closed  systems  con- 
sisting of  those  which  have  a  device  to  prevent  the  return  of  air  after  it  is 
once  expelled  from  the  system,  and  which  can  operate  at  sub-atmospheric 
pressures  for  a  period  of  four  to  eight  hours  depending  upon  the  tightness 
of  the  system,  and  (2)  open  systems  consisting  of  those  which  have  the 
return  line  constantly  open  to  the  atmosphere  without  a  check  or  other 
device  to  prevent  the  return  of  air,  and  which  operate  at  a  few  ounces 
above  atmospheric  pressure. 

508 


CHAPTER  31 — STEAM  HEATING  SYSTEMS 


Under  the  first  classification  the  essentials  are  packless  graduated 
valves  on  the  radiators,  thermostatic  return  traps  on  the  returns,  and 
traps  on  all  drips  unless  they  are  water  sealed.  Such  a  system,  illustrated 
in  Fig.  10,  should  be  equipped  with  an  automatic  return  trap  to  prevent 
the  water  from  backing  out  of  the  boiler.  In  this  up-feed  arrangement 
the  supply  piping  is  carried  to  a  high  point  directly  at  the  boiler  and  is 
graded  down  toward  the  end  or  ends  of  the  supply  main,  each  supply 
main  being  dripped  at  the  end  into  the  wet  return  or  carried  back  to  a 
point  near  the  boiler  where  it  drops  down  below  the  boiler  water  line  and 
becomes  a  wet  return.  From  this  main,  runouts  are  branched  off  to  feed 
risers  or  radiators  above,  these  being  graded  back  toward  the  steam  main 


TRAP 


DRIPPED 

1^ 

BOILER  VATER  LIKE 


FIG.  10.    TYPICAL  UP-FEED  VAPOR  SYSTEM  WITH  AUTOMATIC  RETURN  TRAP* 

^Proper  piping  connections  are  essential  with  special  appliances  for  pressure  equalizing  and  air  elimination. 

if  they  are  not  dripped  at  the  bottom  of  the  riser,  or  toward  the  riser  if 
the  riser  heel  is  dripped.  Both  conditions  are  illustrated  in  Figs.  2  and  3. 
Return  risers  are  connected  to  each  radiator  on  its  return  end  through 
thermostatic  traps.  Their  bottoms  are  connected  to  the  return  main 
through  runouts  which  slope  toward  the  main.  The  return  main  itself  is 
sloped  back  toward  the  boiler  if  it  is  carried  overhead;  if  run  wet,  the 
slope  may  be  neglected.  An  air  vent  is  installed  at  the  point  at  which  the 
return  main  drops  below  the  water  line.  In  the  simplest  cases  this  vent 
consists  of  a  %-in.  pipe  with  a  check  valve  opening  outward,  but  in 
certain  patented  systems  special  forms  of  vent  valves,  designed  to  allow 
the  air  readily  to  pass  out  of  the  system  and  to  prevent  its  return,  are 
used.  A  check  valve  is  inserted  in  the  return  main  at  a  point  near  the 
boiler  and  a  vertical  pipe  is  run  up  into  the  bottom  of  the  return  trap, 
which  usually  is  located  with  the  bottom  about  18  in.  above  the  boiler 
water  line.  Some  traps  are  constructed  to  permit  the  bottom's  being 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


From  steam  main 


ir  vent  and  check 

===== 
Main  return 

•Automatic  return  trap 

sually  18  in. 
Boiler  water  line 


FIG.  11.    TYPICAL  CONNECTIONS  FOR  AUTOMATIC  RETURN  TRAP 

placed  as  close  as  8  in.  above  the  boiler  water  line.  On  the  other  side  of 
this  connection  a  second  check  valve  is  installed  in  the  main  return  just 
before  it  enters  the  boiler  (Fig.  11). 

Down-Feed  Two-Pipe  Vapor  System 

In  the  down-feed  two-pipe  vapor  system  the  steam  is  carried  to  the  top 
of  the  building,  the  top  of  the  vertical  riser  constituting  the  high  point  of 
the  system,  and  the  horizontal  supply  main  is  sloped  down  from  this 
location  to  the  far  ends  of  each  branch.  The  branches  are  taken  off  the 
main  from  the  bottom  or  at  a  45-deg  angle  downward,  with  the  runouts 


Bottom  of 
steam  drop 


Drip  trap 


Graduated  valve 


-Connected  to  dry  return 
(where  connected  to  wet 
return,  drip  trap  may 
be  omitted) 

FIG.  12.    DETAIL  OF  DRIP  CONNECTIONS  AT  BOTTOM  OF  DOWN-FEED  STEAM  DROP 

510 


CHAPTER  31 — STEAM  HEATING  SYSTEMS 


sloped  toward  the  drops  (Fig.  6).  Thus  each  branch  from  the  main  forms 
a  drip  and  no  accumulation  of  water  is  carried  down  any  one  drop. 
Another  method  of  running  the  steam  main,  which  is  not  considered  as 
satisfactory  but  which  is  practical,  is  to  take  the  branches  off  the  top  of 
the  main  (Fig.  7)  and  to  drip  the  end  of  the  main  through  the  last  riser,  as 
illustrated  in  the  down-feed  one-pipe  system  detail  shown  in  Fig.  6.  If 
this  is  done,  the  pipe  drop  at  the  end  or  ends  of  the  mains  should  be 
enlarged  one  pipe  size  to  provide  capacity  for  this  concentration  of  the 
main  drip. 

The  steam  drops  are  carried  down  through  the  building  with  suitable 
reductions  as  the  various  radiator  connections  are  taken  off  until  the 
lowest  radiator  runout  is  reached.  If  the  drop  is  only  two  or  three  stories 
high,  the  portion  feeding  the  bottom  radiator  should  be  increased  one 
pipe  size  to  provide  for  draining  the  riser,  and  if  the  drop  is  over  three 
stories  high  it  is  well  to  increase  the  portion  feeding  the  two  lowest  radi- 
ators one  or  two  pipe  sizes,  especially  if  the  two  lowest  radiators  are  small 
and  the  normal  size  of  drop  required  is  1  in.  or  less.  The  bottom  of  the 
steam  drops  should  terminate  with  a  dirt  pocket  above  which  a  drip  trap 
connection  is  located,  as  shown  in  Fig.  12.  The  returns  on  a  down-feed 
vapor  system  are  the  same  as  on  an  up-feed  system  except  that  every 
steam  drop  must  have  a  drip  at  the  bottom  connected  either  into  the 
return  through  a  trap  or  into  a  separate  water-sealed  drip  line  below  the 
boiler  water  line,  as  illustrated  in  Fig.  10,  in  which  case  the  thermostatic 
traps  may  be  omitted.  The  runouts  to  the  radiators  and  the  radiator 
connections  of  the  down-feed  system  are  the  same  as  those  of  the  up-feed 
system  already  described. 

ATMOSPHERIC  SYSTEM 

The  distinguishing  features  of  the  atmospheric  system  are  gravity 
return  to  the  boiler  or  to  waste,  graduated  or  ordinary  radiator  valves,  no 
automatic  air  valves  on  the  radiators ,  thermostatic  traps  on  the  radiator 
returns,  and  the  venting  of  all  air  from  the  system  by  means  of  pipes  open 
to  the  atmosphere.  The  returns  are  open  to  the  atmosphere  at  all  times, 
usually  by  extending  the  return  risers  to  the  top  of  the  building  where 
they  are  either  connected  together  in  groups  and  carried  through  the  roof 
or  extended  through  the  roof  individually.  Atmospheric  systems,  either 
up-feed  or  down -feed,  are  often  used  where  the  condensation  is  not 
returned  to  the  boiler,  as  in  heating  systems  supplied  by  high  pressure 
steam  through  pressure-reducing  valves  at  locations  far  from  the  boilers. 
The  returns  may  be  delivered  back  to  the  boiler,  if  desired,  by  condensa- 
tion return  pumps  which  are  vented  to  the  atmosphere.  The  return  lines 
in  such  systems  are  simply  gravity  waste  lines  in  which  the  condensation 
flows  entirely  by  gravity  and  is  not  aided  by  any  pressure  difference. 

The  steam  side  may  be  run  as  that  for  either  up-feed  or  down-feed 
two-pipe  vapor  systems,  as  the  conditions  require,  and  the  radiator  con- 
nections are  the  same  as  for  vapor  systems  in  that  they  have  graduated 
valves  on  the  radiator  supply  ends  and  thermostatic  traps  on  the  radiator 
return  ends.  All  drips  from  the  supply  main  and  the  steam  side  of  the 
system  must  pass  through  thermostatic  drip  traps  before  entering  the 
return  system  where  only  atmospheric  pressure  exists.  Fig.  13  illustrates 
a  typical  scheme  of  piping  used  on  atmospheric  systems. 

511 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


Trap 


Trap. 


- 4='~ 


Supply  main, 


^Air  eliminating  and  pressure 


/Trap 


-  Give  good  pitch 


Trap. 


Dry 
returnN 


X  Riser       i    /' 


equalizing  device  j     ;, .    L     ' 

fe*  note  below*  dripped    .£ 


j  See  note  below  * 

^Boiler  water  line    J 


^xWet  return  ^  ^,' 

FIG.  13.   TYPICAL  ATMOSPHERIC  SYSTEM  WITH  AUTOMATIC  RETURN  TRAP* 

«Proper  piping  connections  are  essential  with  special  appliances  for  pressure  equalizing  and  air  elimination 


TRAP- 


HARTFORD 
RETURN 
CONNECTlOh; 


FIG.  14.    TYPICAL  UP-FEED  VACUUM  PUMP  SYSTEM 
512 


CHAPTER  31 — STEAM  HEATING  SYSTEMS 


VACUUM  SYSTEM 

In  the  vacuum  system,  a  vacuum  is  maintained  in  the  return  line 
practically  at  all  times  but  no  vacuum  is  carried  on  the  steam  side,  and  the 
usual  accessories  include  graduated  valves  on  the  radiator  supply  and 
thermostatic  traps  on  the  radiator  return.  The  air  is  expelled  from  the 
system  by  a  vacuum  pump  and  all  drips  must  pass  through  thermostatic 
traps  before  connecting  to  the  return  side  of  the  system. 

These  systems  are  often  fed  from  high  pressure  steam  mains  through 
pressure-reducing  valves  but  they  may  be  fed  direct  from  a  low-pressure 
steam  heating  boiler  as  shown  in  Fig.  14,  in  which  a  typical  up-feed 
vacuum  system  is  illustrated.  The  supply  main  slopes  down  in  the 
direction  of  flow;  the  runouts  pitch  down  toward  the  riser  if  the  riser  is 
dripped  (Fig.  3)  or  up  toward  the  riser  if  the  riser  is  not  dripped  (Fig.  2) ; 
both  conditions  are  indicated  in  Fig.  14.  The  matter  of  dripping  the 
risers  depends  largely  on  the  height  of  the  riser  and  the  judgment  of  the 
designer.  Ordinarily  risers  less  than  three  stories  high  are  not  dripped 
and  those  more  than  four  stories  high  are  dripped,  but  there  is  no  set  rule 
for  this.  When  risers  are  dripped  the  runouts  from  the  steam  main  may 
be  taken  from  the  bottom  if  desired  and  each  runout  then  serves  as  a  drip 
for  the  main. 

The  risers  are  carried  up  to  the  highest  radiator  connection  and  are 
connected  to  the  radiator  through  runouts  sloping  back  toward  the  riser. 
The  radiators  usually  have  graduated  valves  on  the  supply  end,  although 
this  is  not  absolutely  necessary.  Angle-globe  valves  and  gate  valves  may 
be  used  where  graduated  manual  control  is  not  desirable.  The  return 
valves  must  be  of  the  thermostatic  type  which  will  pass  air  and  water  but 
which  will  close  against  the  passage  of  steam. 

The  return  risers  are  carried  down  to  the  basement  and  are  connected 
into  a  common  return  line,  care  being  taken  that  no  air  pockets  exist  in 
the  runouts  or  in  the  horizontal  return  main  which  slopes  downward 
toward  the  vacuum  pump  to  which  it  is  connected.  The  air  and  water 
are  taken  by  the  vacuum  pump,  which  discharges  the  air  from  the  system 
and  pumps  the  water  back  to  the  boiler,  or  other  receiver,  which  may  be  a 
feed-water  tank  or  a  hot  well.  It  is  essential  on  these  systems  that  no 
connection  from  the  supply  side  to  the  return  side  be  made  at  any  point 
except  through  a  trap. 

While  the  best  practice  demands  a  return  flowing  to  the  vacuum  pump 
in  an  interrupted  downward  slope,  in  some  cases  limitations  make  it 
necessary  to  drop  the  return  below  the  level  of  the  vacuum  pump  inlet 
before  the  pump  can  be  reached.  In  such  event  one  of  the  advantages  of 
the  vacuum  system  is  that  the  return  can  be  raised  by  the  suction  of  the 
vacuum  pump  to  a  considerable  height,  depending  on  the  amount  of 
vacuum  maintained,  by  means  of  a  lift  fitting  inserted  in  the  return. 
When  the  lift  is  considerable,  several  lift  fittings  are  used  in  steps  (Fig. 
15),  more  successful  operation  being  obtained  by  this  method  than  when 
the  lift  is  made  in  one  step.  If  the  lift  occurs  close  to  the  vacuum  pump, 
a  special  arrangement  is  used  as  shown  in  Fig.  16. 

Down -Feed  Vacuum  System 

The  piping  arrangement  for  the  down-feed  vacuum  system  is  similar 
on  the  supply  side  to  the  down -feed  vapor  system  in  that  it  has  similar 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


runouts,  radiator  valves,  drips  on  the  bottom  of  the  steam  drops,  and 
enlargement  of  the  drops  for  the  lower  radiator  connections.  The  return 
side  of  the  system  is  exactly  the  same  as  the  up-feed  system  except  that 
the  steam  riser  drips  at  the  bottom  are  connected  into  the  return  line 
through  thermostatic  traps.  It  is  preferable  to  take  the  runouts  for  the 
risers  from  the  bottom  or  at  a  45-deg  angle  down  from  the  steam  main 
(Fig.  6)  so  that  they  may  serve  as  steam  main  drips.  When  this  is  done 
it  is  practical  to  run  the  steam  main  level  if  a  runout  is  located  at  every 
change  in  pipe  size,  or  if  eccentric  fittings  are  used  (Fig.  17).  A  slight 
pitch  in  the  steam  main,  however,  should  be  used  when  possible.  An 
overhead  vacuum  down-feed  system  is  shown  diagrammatically  in  Fig.  18. 


CLOSE  NIPPLE 


VERTICAL  LIFT  TO 
BE  ONE  SIZE.   .• 
SMALLEC  THAN  THE 
VACUVM 


CUFT  FITTING 
^WvCUUM  2ETUQN 


U-UFT  FITTING 


VACUUM  CETUBN 
./LIFT  FITTING 


FIG.  15.     METHOD  OF  MAKING  LIFTS 

ON  VACUUM  SYSTEMS  WHEN  DISTANCE 

is  OVER  5  FT 


FIG.  16.     DETAIL  OF  MAIN  RETURN 
LIFT  AT  VACUUM  PUMP 


ECCENTRIC  QEDUONG 
(COUPLING. 


FIG.  17.    METHOD  OF  CHANGING  SIZE  OF  STEAM  MAIN  WHEN  RUNOUTS 
ARE  TAKEN  FROM  TOP 


SUB-ATMOSPHERIC  SYSTEMS 

The  sub-atmospheric  systems  are  similar  to  the  vacuum  system  except 
that  a  pump  capable  of  operating  up  to  25  in.  of  vacuum  is  used,  and  a 
control  is  placed  on  the  pump  so  that  the  vacuum  or  absolute  pressure 
carried  in  the  return  can  be  maintained  a  certain  amount  below  that 
existing  in  the  steam  line  to  cause  a  constant  circulation.  The  traps  are 
designed  to  operate  in  high  vacuum.  It  is  apparent  that  this  system 
differs  from  the  ordinary  vacuum  system  by  having  a  vacuum  on  both 
sides  of  the  system,  instead  of  only  on  the  return  side,  in  order  to  secure 
control  of  the  heat  emission  from  the  radiators  and  thus  to  control  the 
temperature  in  the  building.  The  system  can  be  operated  in  the  same 
manner  as  the  ordinary  vacuum  system  when  desired. 

In  the  vacuum  system,  steam  pressure  above  that  of  the  atmosphere 
exists  in  the  supply  mains  and  radiators  practically  at  all  times.  In  the 
sub-atmospheric  system,  steam  pressure  exists  in  the  steam  main  and 

514 


CHAPTER  31 — STEAM  HEATING  SYSTEMS 


radiators  only  during  the  most  severe  weather,  while  under  average 
winter  temperatures  the  steam  is  under  a  partial  vacuum  which  in  mild 
weather  may  reach  as  high  as  25  in.  This  vacuum  is  largely  self-induced 
by  the  condensation  of  the  steam  in  the  system  when  an  inadequate 
supply  of  steam  is  being  furnished  through  the  control  valve  which  admits 
it.  In  the  sub-atmospheric  system,  a  control  valve  is  inserted  on  the 
steam  main  of  an  ordinary  vacuum  system  near  the  boiler,  a  high-vacuum 
pump  is  substituted  for  the  ordinary  type  and  is  supplied  with  a  pressure- 
difference  control,  and  traps  are  placed  on  the  radiators  and  drips  which 
will  operate  satisfactorily  at  any  pressure  from  5  Ib  gage  to  26  in.  of 
vacuum. 


Pitch 


Loop  /  Vacuum  pump 

-fir 

FIG.  18.    TYPICAL  DOWN-FEED  VACUUM  SYSTEM 

The  control  valve  is  a  special  pressure-reducing  valve  which  may  be 
controlled  manually  or  thermostatically  from  points  selected  in  the 
building.  The  vacuum  pump  regulator  is  simply  a  diaphragm  so  ar- 
ranged that,  when  the  vacuum  in  the  return  line  is  insufficient  to  hold  the 
desired  difference  in  pressure  between  the  steam  and  return  sides  of  the 
system,  the  vacuum  pump  is  automatically  started  and  the  vacuum 
increased  to  the  necessary  amount.  The  actual  pressure  difference  main- 
tained between  the  two  sides  of  the  system  is  only  enough  to  secure 
adequate  circulation  and  is  often  about  2  in.  of  mercury.  This  fixed 
pressure  difference  between  the  supply  and  return  sides  of  the  system 
results  in  practically  constant  circulation  under  all  pressure  conditions. 

In  order  to  distribute  the  steam  equally  when  the  system  is  being 
warmed  up  and  also  to  reduce  the  amount  of  steam  delivered  to  the 
radiators  on  mild  days,  orifice  plates  are  used  in  the  graduated  radiator 
control  valves.  The  heat  emitted  from  the  radiators  in  mild  weather  and 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

under  conditions  of  high  vacuum  is  not  only  reduced  in  proportion  to  the 
difference  in  the  steam  temperature  between  that  for  2  Ib  gage  and  for 
25  in.  of  vacuum  but  it  is  reduced  still  further  by  a  reduction  in  the  amount 
of  steam  which  can  pass  through  the  orifice  when  the  steam  is  expanded 
due  to  the  vacuum.  This  renders  possible  the  control  of  heat  emission 
from  the  radiators  to  a  point  not  indicated  entirely  by  the  difference  in 
steam  temperatures. 

The  high-vacuum  pumps  on  this  system  are  equipped  with  receivers 
having  float  control  so  that  the  pump  can  be  placed  on  a  receiver-return- 
pump  basis  at  night  if  desired  so  no  high  vacuum  will  be  carried.  One 
radical  difference  between  this  system  and  the  ordinary  vacuum  system 
is  that  no  lifts  can  be  made  in  the  return  line.  The  returns  must  grade 
downward  constantly  and  uninterruptedly  from  the  radiator  return 
outlet  to  the  inlet  on  the  high-vacuum  pump  receiver.  No  attempt  should 
be  made  to  heat  service  water  on  this  system  unless  the  steam  line  for 
water  heating  is  taken  off  the  boiler  header  back  of  the  heating  system 
control  valve,  and  then  only  when  2  Ib  or  more  will  be  carried  on  the 
boiler  at  all  times. 

ORIFICE  SYSTEM 

Orifice  systems  of  steam  heating  may  have  piping  arrangements 
identical  with  vacuum  systems  but  some  of  these  systems  omit  both  the 
radiator  thermostatic  traps  and  the  vacuum  pump  in  cases  where  the 
returns  are  wasted  to  a  sewer  or  delivered  to  some  type  of  receiver  in 
which  no  back  pressure  exists.  The  principle  on  which  they  operate  is 
embodied  in  the  well-known  fact  that  an  orifice  will  deliver  varying 
velocities  when  the  ratio  of  the  absolute  pressures  on  the  two  sides  of  the 
orifice  exceeds  58  per  cent.  If  the  absolute  pressure  on  the  outlet  side  is 
less  than  58  per  cent  of  the  absolute  pressure  on  the  inlet  side  no  further 
increase  in  velocity  will  be  obtained. 

As  a  result,  if  an  orifice  is  so  designed  in  size  as  to  exactly  fill  a  radiator 
with  steam  at  2-lb  gage  on  one  side  and  J^-lb  gage  on  the  other,  the  abso- 
lute pressure  relation  is 

14.7  +  0.25        on 

r-  —  90  per  cent 


14.7  + 

Should  the  steam  pressure  be  dropped  to  %  Ib  gage,  the  pressure  on  each 
side  of  the  orifice  would  be  balanced  and  no  steam  flow  would  take  place. 
From  this  it  will  be  seen  that  if  an  orifice  of  a  given  diameter  will  fill  a 
given  radiator  with  steam  when  there  is  a  given  pressure  on  the  main,  it  is 
simply  a  question  of  dropping  this  main  pressure  so  as  to  fill  any  desired 
portion  of  the  radiator  down  to  the  point  where-  the  main  pressure  equals 
the  back  pressure  in  the  radiator,  at  which  time  no  steam  will  be  supplied 
at  all.  If  orifices  throughout  a  job  are  designed  on  a  similar  basis,  all 
radiators  will  heat  proportionately  to  the  steam  pressure  within  the  limits 
for  which  the  orifices  are  designed. 

Some  systems  use  orifices  not  only  in  radiator  inlets  but  also  at  different 
points  on  the  main,  thus  balancing  the  system  to  a  greater  extent.  For 
example,  the  system  may  be  designed  for  a  particularly  long  run  involving 
an  initial  pressure  of  3-lb  gage  on  the  main  and  2  Ib  at  the  end  of  the  main, 

516 


CHAPTER  31 — STEAM  HEATING  SYSTEMS 


but  each  branch  from  the  main  may  have  an  orifice  for  reducing  the 
pressure  at  it  to  2  Ib-gage.  This  is  particularly  useful  for  branches  near 
the  boiler  where  the  drop  in  the  main  has  not  yet  been  produced. 

Orifice  systems  using  a  vacuum  pump  operate  successfully  with  the 
ordinary  low  vacuum  type  of  pump  producing  8  to  10  in.  of  vacuum. 
They  are  controlled  by  various  means  to  regulate  the  steam  pressure. 
One  method  is  by  a  thermostat  located  on  the  roof  to  govern  the  steam 
pressure  by  a  combination  of  outside  and  inside  temperatures;  another, 
useful  on  systems  without  traps  and  vacuum  pumps,  controls  the  steam 
pressure  manually  from  temperature  indication  stations  in  the  building, 
or  automatically  by  a  thermostatically-controlled  pressure  reduction 
valve  or  draft  regulator  on  the  boiler ;  with  oil  or  gas  firing,  the  on-and-off 
control  or  a  boiler  pressure  control  may  be  used. 

ZONE  CONTROL 

Certain  portions  of  a  building  may  require  more  heat  at  times  than 
others  but  if  the  whole  building  is  on  one  general  control,  such  as  would 
occur  with  a  single  piping  system  with  an  on-and-off  control  or  with  the 
sub-atmospheric  or  the  orifice  systems,  it  would  be  necessary  to  supply 
sufficient  heat  to  accommodate  the  coldest  portion  of  the  building  even 
though  some  sections  would  be  overheated.  By  zoning,  each  section  of  a 
building  may  be  controlled  separately. 

The  sides  of  the  building  with  different  exposures  should  be  considered 
first,  because  of  the  varying  effects  of  the  wind  and  sun.  With  the  pre- 
vailing winter  winds  from  the  northwest,  a  simple  zoning  would  place  the 
north  and  west  sides  of  the  building  on  one  system  and  the  south  and  east 
sides  on  another.  If  the  building  is  large  enough  to  justify  the  expendi- 
ture, a  better  arrangement  would  be  to  place  all  north  walls  on  one  zone, 
all  west  walls  on  a  second,  all  east  walls  on  a  third,  and  all  south  walls  on 
a  fourth. 

In  case  of  high  buildings,  the  lowest  8  or  10  stories  may  be  well  protected 
from  wind  by  surrounding  buildings,  the  next  10  stories  may  have 
moderate  exposure,  and  above  this  there  may  be  an  unobstructed  exposure 
to  gales.  On  still  days  the  heat  demands  vertically  will  vary  little,  but  on 
windy  days  there  will  be  a  marked  difference  in  the  heat  requirements  for 
the  different  horizontal  sections.  In  addition,  the  chimney  effect  caused 
by  the  difference  in  density  between  the  warm  air  on  the  inside  of  a 
building  and  the  colder  air  on  the  outside  will  give  an  air  movement  which 
will  require  zoning  to  correct.  Where  such  conditions  are  encountered, 
the  building  should  be  divided  horizontally  as  well  as  vertically.  An 
arrangement  of  this  character  would  give  12  zones:  namely,  north,  east, 
south,  and  west  lower  zones;  similar  middle  zones;  and  similar  top  zones. 
Each  zone  should  constitute  an  individual  and  separate  system  of  piping 
with  its  own  supply  steam  valve  (controlled  by  thermostats  in  its  respec- 
tive zone)  and  with  its  own  return  or  vacuum  pump,  if  one  is  used- 
Certain  interior  areas,  such  as  basements,  light  well  walls  and  other 
locations  where  sun  and  wind  do  not  affect  the  conditions,  should  be 
placed  in  still  another  zone  if  the  most  economical  results  are  to  be 
secured. 

Zoning  has  advantages  even  where  individual  thermostatic  radiator 

517 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

control  is  installed  whether  this  be  of  pneumatic,  electric,  or  the  self- 
contained  radiator  valve  type.  By  operating  the  different  zones  to 
parallel  outside  temperature  requirements,  a  large  part  of  the  load  is 
taken  off  the  thermostatic  controls ;  they  make  fewer  operations  and  the 
radiator  follows  a  more  even  temperature  instead  of  fluctuating  from 
extreme  hot  to  extreme  cold. 

CONDENSATION  RETURN  PUMPS 

Condensation  return  pumps  are  generally  required  when  the  elevation 
of  the  boiler  with  respect  to  the  heating  units  is  such  that  the  condensate 
will  not  return  by  gravity,  or  when  the  boiler  pressure  is  greater  than  that 


TPAP- 


SUPPLY  MAINi 


DRY  RETURN" 
f-AUTOMWTC  WTER  FEEDER 


-AIR  VENT. 
-AUTOMATIC  PUrtP  8c  RECEIVER 

^eYttVSS  TO  DRAIN 
FIG.  19.    TYPICAL  INSTALLATION  USING  CONDENSATION  PUMP 

supplied  the  heating  units,  as  in  a  high-pressure  boiler  installation  sup- 
plying steam  through  a  reducing  valve  to  the  heating  units.  The  con- 
densate is  commonly  returned  by  gravity  to  a  receiver,  vented  to  the 
atmosphere,  from  which  it  flows  to  the  pump. 

Condensation  return  pumps  are  assembled  with  tank  or  receiver  and 
arranged  for  either  continuous  operation  or  for  automatic  starting  and 
stopping  by  float  control.  Any  style  of  water  pump  may  be  employed  for 
this  service,  the  power  available  determining  whether  the  mode  of  drive 
snail  be  steam  or  electric.  The  motor-driven,  automatic,  centrifugal 
pump  and  receiver  has  found  wide  acceptance  in  practice  for  low  pressure 
heating  systems. 

Fig.  19  shows  a  typical  installation  using  an  automatic  condensation 
return  pump  and  vented  receiver,  A  float  control  operates  the  pump 
whenever  sufficient  water  accumulates.  Condensation  return  pumps  are 

518 


CHAPTER  31 — STEAM  HEATING  SYSTEMS 


suitable  for  use  on  systems  in  which  the  returns  are  under  atmospheric 
pressure.  These  include  atmospheric  systems,  orifice  systems  with  open 
returns,  and  certain  types  of  vapor  systems  which  operate  within  a  few 
ounces  of  atmospheric  pressure,  but  ordinarily  do  not  carry  any  sub- 
atmospheric  pressure.  They  may  also  be  used  on  one-pipe  and  two-pipe 
gravity  steam  systems  with  a  proper  arrangement  for  venting  the  receiver. 
In  discharging  to  waste  there  is  no  object  in  using  a  condensation  pump 
unless  the  discharge  mr.ot  be  elevated. 

VACUUM  PUMPS 

A  vacuum  heating  pump  is  employed  to  create  a  vacuum  on  the  return 
end  of  a  system  to  remove  air  and  water  and  to  return  the  condensate  to 
the  boiler  or  to  some  other  intercepting  device  that  may  be  employed  in 
plants  having  mixed  systems  of  heating  and  other  services.  Pumps  of 
this  classification  may  be  driven  by  steam  or  electricity;  they  may  be 
continuous  in  operation,  or  automatic  with  float  or  vacuum  control  in  one 
or  more  combinations. 

For  rating  purposes3,  vacuum  pumps  are  classified  as  low  vacuum  and 
high  vacuum.  Low  vacuum  pumps  are  those  rated  under  operation  at 
5^-in.  mercury  vacuum,  and  high  vacuum  pumps  are  those  rated  at 
vacuums  above  5J^  in. 

Return  line  vacuum  pumps  are  classified  in  the  method  of  their  per- 
formance as  follows: 

a.  Those  which  perform  the  function  of  air  separation  under  atmospheric  pressure. 

b.  Those  which  perform  the  function  of  air  separation  under  a  partial  vacuum. 

Pumps  coming  under  the  first  classification  will  handle  vacuum  steam 
system  condensation  coming  back  by  gravity  at  any  temperature  up  to 
205  F  without  either  the  sealing  or  the  hurling  water  flashing  into  steam. 
These  pumps,  to  operate  under  a  combined  water  level  and  vacuum  con- 
trol, must  be  equipped  with  a  float-control  receiver  between  the  vacuum 
pump  and  the  system,  but  where  they  are  intended  for  continuous  opera- 
tion, they  do  not  require  a  receiver.  Such  pumps  employ  a  single  vacuum 
producer  which  removes  the  condensate  and  air  from  the  system  and 
delivers  it  into  a  separating  chamber  under  atmospheric  pressure  from 
which  the  condensate  is  delivered  to  the  boiler  or  feed  water  heater.  They 
are  constructed  on  one  of  the  following  evacuating  and  discharge  principles : 

1.  Hydraulic  vacuum  producer  with  one  pump  impeller. 

2.  Hydraulic  vacuum  producer  with  two  pump  impellers. 

3.  Water  displacement  vacuum  producer  with  two  pump  impellers. 

4.  Piston  displacement  vacuum  producer  with  one  pump  piston. 

The  second  classification  of  pumps  will  handle  vacuum  steam  system 
condensation  coming  back  by  gravity  at  any  temperature  not  exceeding 
190  F  without  the  flashing  into  steam  of  either  the  sealing  or  the  hurling 
water.  In  order  to  operate  under  a  combined  water-level  and  vacuum 
control,  these  pumps  must  be  equipped  with  a  float-control  receiver 
between  the  vacuum  pump  and  the  system;  where  intended  for  con- 
tinuous operation  they  do  not  require  a  receiver.  Such  pumps  employ  a 
vacuum  producing  impeller  which  removes  air  from  the  receiver  or 


3See  A.S.H.V.E,  Standard  Code  for  Testing  and  Rating  return  line  low  vacuum  heating  pump. 

519 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

heating  system  under  a  partial  vacuum  and  delivers  it  through  an  air 
separator  against  atmospheric  pressure.  The  condensate  is  removed 
from  the  receiver  under  a  partial  vacuum  by  a  separate  impeller  and  is 
delivered  to  the  boiler  or  feed  water  heater.  For  evacuating  and  dis- 
charge, a  water  displacement  vacuum  producer  with  two  pump  impellers 
is  used. 

Receiver  Capacities  for  Vacuum  Pumps 

Where  receivers  are  used  in  connection  with  vacuum  pumps  there  is  a 
definite  relation  between  the  capacity  of  the  receiver  and  the  capacity 
of  the  pump.  The  receiver  should  have  a  capacity  of  not  less  than  1J^ 
times  the  volumetric  quantity  of  condensation  per  minute  and  should  not 
have  such  a  capacity  that  the  pump  will  empty  the  receiver  in  less  than 
half  a  minute.  Receivers  of  larger  capacities  will  result  in  less  frequent 
periods  of  operation. 

Piston  Displacement  Vacuum  Pumps 

Piston  displacement  return-line  vacuum  heating  pumps  may  be  either 
power  or  steam  driven.  They  should  be  provided  with  mechanical 
lubricators  and  their  piston  speed  in  feet  per  minute  should  not  exceed 
20  times  the  square  root  of  the  number  of  inches  in  their  stroke.  While 
the  volumetric  displacement  for  such  pumps  was  formerly  figured  at  8  to 
10  times  the  volumetric  flow  of  condensation  to  be  handled,  the  more 
efficient  thermostatic  traps  used  today  in  connection  with  vacuum 
heating  systems  make  it  possible  to  change  this  proportion  so  that  the 
volumetric  displacement  of  these  pumps  may  not  be  less  than  6  times  the 
volume  of  condensation. 

Vacuum  Pump  Controls 

In  the  ordinary  vacuum  system  the  vacuum  pump  is  controlled  by  a 
vacuum  regulator  which  cuts  in  when  the  vacuum  drops  to  the  lowest 
point  desired  and  which  cuts  out  when  the  vacuum  has  been  increased  to 
the  highest  point.  This  is  done  largely  to  eliminate  the  constant  starting 
and  stopping  of  the  vacuum  pump  which  would  occur  if  the  vacuum  were 
maintained  constant.  In  addition  to  this  control,  a  float  control  is  in- 
cluded which  will  automatically  start  the  pump  whenever  sufficient  con- 
densation accumulates  in  the  receiver,  regardless  of  the  vacuum  in  the 
system.  This  arrangement  makes  the  vacuum  pump  primarily  a  con- 
densation pump  and  secondarily  an  air  pump. 

On  the  sub-atmospheric  systems  the  high  vacuum  pump  is  controlled 
by  a  differential  regulator  which  keeps  the  vacuum  in  the  return  line 
always  a  few  inches  higher  than  that  in  the  steam  line  and  in  the  radiators. 

TRAPS 

Traps  are  used  for  draining  the  condensate  from  radiators,  steam 
piping  systems,  kitchen  equipment,  laundry  equipment,  hospital  equip- 
ment, drying  equipment  and  many  other  kinds  of  apparatus.  The  usual 
functions  of  a  trap  are  to  allow  the  passage  of  condensate  and  to  prevent  the 
passage  of  steam.  In  addition  to  these  functions,  traps  are  frequently 
required  to  allow  the  passage  of  air  as  well  as  condensate.  Traps  are  also 

520 


CHAPTER  31 — STEAM  HEATING  SYSTEMS 


required  to  allow  the  passage  of  air  and  to  prevent  the  passage  of  either 
water  or  steam,  or  both. 

In  addition,  traps  are  used  for  returning  condensate  either  by  gravity, 
by  steam  pressure,  or  by  both,  to  a  boiler  or  other  point  of  disposal,  and 
for  lifting  condensate  from  a  lower  to  a  higher  elevation,  or  for  handling 
condensate  from  a  lower  to  a  higher  pressure. 

The  fundamental  principle  upon  which  the  operation  of  practically  all 
traps  depends  is  that  the  pressure  within  the  trap  at  the  time  of  discharge 
shall  be  equal  to,  or  slightly  in  excess  of,  the  pressure  against  which  the 
trap  must  discharge,  including  the  friction  head,  velocity  head  and  static 
head  on  the  discharge  side  of  the  trap.  If  the  static  head  is  in  favor  of 
the  trap  discharge  it  is  a  minus  quantity  and  may  be  deducted  from  the 
other  factors  of  the  discharge  head. 

Traps  may  be  classified  according  to  the  principle  of  operation  as  (1) 
float,  (2)  bucket,  (3)  thermostatic,  or  (4)  tilting  traps. 

Float  Traps.  A  discharge  valve  is  operated  by  the  rise  and  fall  of  a  float  due  to  the 
change  of  water  level  in  the  trap.  When  the  trap  is  empty  the  float  is  in  its  lowest 
position,  and  the  discharge  valve  is  closed.  A  gage  glass  indicates  the  height  of  water 
in  the  chamber. 

Unless  float  traps  are  well  made  and  proportioned  there  is  danger  of  considerable 
steam  leakage  through  the  discharge  valve  due  to  unequal  expansion  of  the  valve  and 
seat  and  the  sticking  of  moving  parts.  The  discharge  from  a  float  trap  is  usually  con- 
tinuous since  the  height  of  the  float,  and  consequently  the  area  of  the  outlet,  is  propor- 
tional to  the  amount  of  water  present. 

Bucket  Traps.  Bucket  traps  are  of  two  types,  the  upright  and  inverted,  and  although 
they  are  both  of  the  open  float  construction,  their  operating  principle  is  entirely  different. 
In  the  upright  bucket  trap,  the  water  of  condensation  enters  the  trap  and  fills  the  space 
between  the  bucket  and  the  walls  of  the  trap.  This  causes  the  bucket  to  float  and  forces 
the  valve  against  its  seat,  the  valve  and  its  stem  usually  being  fastened  to  the  bucket. 
When  the  water  rises  above  the  edges  of  the  bucket  it  flows  into  it  and  causes  it  to  sink, 
thereby  withdrawing  the  valve  from  its  seat.  This  permits  the  steam  pressure  acting 
on  the  surface  of  the  water  in  the  bucket  to  force  the  water  to  a  discharge  opening.  When 
the  bucket  is  emptied  it  rises  and  closes  the  valve  and  another  cycle  begins.  The  discharge 
from  this  type  of  trap  is  intermittent. 

In  the  inverted  bucket  trap,  steam  floats  the  inverted  submerged  bucket  and  closes  the 
valve.  Water  entering  the  trap  fills  the  bucket  which  sinks  and  through  compound 
leverage  opens  the  valve,  and  the  trap  discharges.  It  is  impossible  to  install  a  water 
gage  glass  on  an  inverted  bucket  trap,  but  if  visual  inspection  is  necessary,  a  gage  glass 
can  be  placed  on  the  line  leading  to  the  trap.  No  air  relief  cocks  can  be  used,  but  this  is 
unnecessary,  as  the  elimination  of  air  is  automatically  taken  care  of  by  air  passing  through 
the  vent  in  the  top  of  the  inverted  bucket  regardless  of  temperature. 

Thermostatic  Traps.  Thermostatic  traps  are  of  two  types,  those  in  which  the  discharge 
valve  is  operated  by  the  relative  expansion  of  metals,  and  those  in  which  the  action  of 
a  volatile  liquid  is  utilized  for  this  purpose.  Thermostatic  traps  of  large  capacity  for 
draining  blast  coils  or  very  large  radiators  are  called  blast  traps. 

Tilting  Traps.  With  this  type  of  trap,  water  enters  a  bowl  and  rises  until  its  weight 
overbalances  that  of  a  counter-weight,  and  the  bowl  sinks  to  the  bottom.  As  the  bowl 
sinks,  a  valve  is  opened  thus  admitting  live  steam  pressure  on  the  surface  of  the  water 
and  the  trap  then  discharges.  After  the  water  is  discharged,  the  counter- weight  sinks 
and  raises  the  bowl,  which  in  turn  closes  the  valve  and  the  cycle  begins  again*  Tilting 
traps  are  necessarily  intermittent  in  operation.  They  are  not  ordinarily  equipped  with 
glass  water  gages,  as  the  action  of  the  trap  shows  when  it  is  filling  or  emptying.  The  air 
relief  of  tilting  traps  is  taken  care  of  by  the  valves  of  the  trap. 

Thermostatic  traps  are  generally  used  for  draining  radiators  and 
heaters,  except  for  very  large  capacities  where  bucket,  float  or  blast-type 
thermostatic  traps  are  used.  Thermostatic  traps  for  this  service  usually 

521 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


pass  both  condensate  and  air  and  in  the  case  of  float  and  upright  bucket 
traps  the  air  is  usually  relieved  through  an  auxiliary  thermostatic  trap  in 
a  by-pass  around  the  main  trap.  Sometimes  this  auxiliary  air  trap  is  an 
integral  part  of  the  trap. 

Blast-type  thermostatic  traps  are  sometimes  used  on  vacuum  heating 
systems  for  connecting  old  one-  or  two-pipe  gravity  systems  in  parallel 
with  vacuum  return  line  systems,  in  which  case  the  blast-type  thermo- 
static traps  should  not  be  provided  with  auxiliary  air  by-pass,  as  the 
action  of  this  will  allow  the  vacuum  to  draw  air  into  the  old  system 
through  its  air  valves,  especially  when  the  steam  is  wholly  or  partially 


Connection  io  Main 
Vacuum  Return 


FIG.  20. 


High  Press 
*    Trap 


METHOD  OF  DISCHARGING  HIGH-PRESSURE  APPARATUS  INTO  LOW-PRESSURE 
HEATING  MAINS  AND  VACUUM  RETURN  MAINS  THROUGH 
A  LOW-PRESSURE  TRAP 


cut  off.  The  air  from  the  returns  of  such  old  systems  should  be  relieved 
just  ahead  of  the  traps  by  means  of  quick-venting  automatic  air  valves, 
preferably  of  the  non-return  type,  especially  if  the  other  air  valves  on 
the  old  system  are  non-return  valves. 

Tilting  traps  used  for  discharging  to  a  higher  or  a  lower  pressure  are 
provided  with  two  or  three  valves  operated  by  the  action  of  the  trap. 
In  the  case  of  the  two-valve  tilting  traps,  one  valve  closes  a  steam  inlet 
and  the  other  valve  opens  a  vent  outlet  while  the  trap  is  filling,  and  as 
soon  as  the  trap  dumps,  the  first  valve  opens  the  steam  inlet  and  the 
second  valve  closes  the  vent  outlet,  while  the  trap  discharges.  In  this 
type  of  trap  there  must  be  a  swinging  check-valve  on  each  side  of  the 
trap,  in  addition  to  the  usual  by-pass,  to  prevent  the  pressure  in  the  trap, 
while  discharging,  from  backing  up  through  the  inlet  and  the  pressure 
in  the  discharge  line  from  backing  up  into  the  trap  while  it  is  filling.  This 
type  of  trap  will  blow  steam  out  through  the  vent  while  filling,  if  the 
pressure  on  the  inlet  side  is  sufficient,  and  should  not  be  used,  therefore, 
with  such  pressures  unless  the  vent  is  properly  piped  back  into  the  return 
to  a  feed  water  heater,  a  condenser  or  a  perforated  pipe  in  the  bottom 
of  the  receiver  to  which  the  trap  discharges  in  such  a  way  as  to  prevent 
the  escape  of  the  steam  that  comes  in  with  the  condensate  and  passes 
through  the  vent.  In  the  three-valve  traps  of  this  type  there  is  an  extra 
valve  for  closing  the  discharge  while  the  trap  is  filling. 

High  pressure  traps  should  not  discharge  directly  into  a  vacuum  return 

522 


CHAPTER  31 — STEAM  HEATING  SYSTEMS 


Returns-., 
Check  Valve,      Tee., 


Gorge, 
Glass 


•C  dir  Valvt 
•Safety  Valve  -h  Wvs+e 

Connection  for 


FIG.  21.     RETURN  TRAP  AND  RECEIVER  FOR  AUTOMATIC  BOILER  FEED 

because  of  the  vapor  formed  by  the  re-evaporation  of  a  part  of  the  hot 
condensation.  Fig.  20  shows  a  method  which  may  be  used  for  disposing  of 
the  greater  part  of  the  vapor  of  re-evaporation. 

Automatic  Return  Traps 

In  the  general  heating  plant,  where  thermostatic  traps  are  installed  on 
the  heating  units,  it  becomes  necessary  to  provide  a  means  for  returning 
the  water  of  condensation  to  the  boiler,  if  a  condensation  or  vacuum  pump 
is  not  used.  When  the  return  main  can  be  kept  sufficiently  high  above  the 
boiler  water  line  for  all  operating  conditions,  the  water  of  condensation 

523 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

will  flow  back  by  gravity,  and  no  mechanical  device  is  required.  But 
actually  this  does  not  work  out  in  practice.  It  follows,  therefore,  that  a 
direct  return  trap  is  needed  for  the  handling  of  the  condensation  even 
though  it  may  not  be  called  into  action  except  under  some  operating 
condition  where  the  pressure  differential  exceeds  the  static  head  provided. 
The  installation  of  a  direct  return  trap  assures  safety  for  such  systems, 
and  guarantees  the  operation  of  the  plant  under  varying  conditions. 

Automatic  return  traps,  sometimes  called  alternating  receivers,  may 
be  of  the  counterbalanced,  tilting  type,  or  spring  actuated.  These  consist 
of  a  small  receiver  with  an  internal  float,  and  when  the  condensate  will 
not  flow  into  the  boiler  under  pressure,  it  will  feed  into  the  receiver  of  the 
trap,  and  in  so  doing,  raise  or  tilt  the  float  or  mechanism  which  actuates  a 
steam  valve  automatically.  This  admits  steam  to  the  receiver,  at  boiler 
pressure,  and  the  equalizing  of  the  pressures  which  follows  allows  the 
water  to  flow  into  the  boiler.  Fig.  21  shows  a  direct  return  tilting  trap 
and  receiver  properly  connected  for  automatically  feeding  a  boiler  from  a 
system  of  returns  delivering  the  condensate  to  the  receiver. 


PROBLEMS  IN  PRACTICE 

1  •  To  what  main  features  does  the  gravity  one-pipe  steam  system  owe  its 
popularity? 

To  its  low  cost  of  installation  and  to  its  simplicity. 

2  •  How  many  types  of  common  mechanical  returns  are  there  and  what  are 
they? 

Three:  (1)  the  mechanical  return  trap,  (2)  the  condensation  return  pump,  and  (3) 
the  vacuum  pump. 

3  •  In  the  ordinary  vacuum  system  of  steam  heating,  where  does  ihe  vacuum 
usually  exist? 

On  the  return  side  of  the  system  only,  between  the  radiator  trap  and  the  vacuum  pump. 
If  the  radiator  supply  valve  is  closed  off,  the  vacuum  may  extend  back  through  the 
radiator  as  far  as  the  supply  valve;  if  an  adequate  supply  of  steam  is  furnished  to  the 
system,  some  vacuum  may  be  developed  in  the  steam  main,  but  neither  of  these  can  be 
termed  normal  operation. 

4  •  What  is  the  distinction  between  the  open  and  the  closed  vapor  systems? 

The  open  vapor  system  has  the  return  line  always  open  to  the  atmosphere,  while  the 
closed  vapor  system  has  an  automatic  device  on  the  air  vent  so  that  air  once  expelled 
from  the  system  through  the  vent  cannot  re-enter  via  this  route. 

5  •  On  a  vacuum  system,  what  device  must  be  placed  on  all  drips  before  they 
enter  the  vacuum  return  line? 

A  thermostatlc  drip  trap  or  occasionally,  where  large  volumes  of  condensation  are  to  be 
handled,  a  float  trap. 

6  •  How  does  the  sub-atmospheric  system  differ  in  operation  from  the  ordinary 
vacuum  system? 

The  ordinary  vacuum  system  has  pressure  in  the  steam  line,  and  a  vacuum  produced  by 
the  vacuum  pump  in  the  return  line,  usually  varying  between  5  and  10  in.  of  water.  The 
sub-atmospheric  system  may  have  either  a  vacuum  or  pressure  on  the  steam  and  return 

524 


CHAPTER  31 — STEAM  HEATING  SYSTEMS 


lines,  but  a  constant  difference  in  pressure  is  maintained  between  the  lines  regardless  of 
what  pressure  or  vacuum  may  be  carried.  The  vacuum,  which  is  generally  produced 
by  condensation  in  the  system  under  conditions  of  throttled  steam  supply,  may  run 
much  higher  than  in  the  ordinary  vacuum  systems. 

7  •  What  is  generally  understood  by  zoning  in  building  steam  heating  systems? 

Zoning  is  a  term  applied  to  the  placing  of  certain  sections  of  a  building  on  a  single 
temperature  control  instead  of  having  either  individual  room  control  or  a  single  tempera- 
ture control  governing  the  whole  building.  Zones  may  be  horizontal,  such  as  a  single 
story,  a  basement,  or  an  attic,  or  vertical  such  as  the  north  side,  or  the  west  side. 

8  •  Why  does  the  water  line  in  the  far  end  of  a  wet  return  in  a  gravity  steam 
system  rise  higher  than  the  water  line  in  the  boiler? 

The  friction  of  the  steam  flowing  through  the  steam  main  from  the  boiler  to  the  far 
end  of  the  system  causes  a  drop  in  steam  pressure  at  the  point  where  the  wet  return  is 
connected ;  consequently,  the  steam  pressure  on  top  of  the  water  in  the  wet  return  is  less 
than  the  steam  pressure  on  top  of  the  water  in  the  boiler,  so  the  water  in  the  end  of  the 
wet  return  rises  until  a  balanced  condition  is  set  up. 

9  •  On  gravity  one-pipe  systems  as  indicated  in  Fig.  1  and  Fig.  3,  why  is  the 
drip  on  the  steam  runout  connected  to  wet  return? 

Because  if  it  were  connected  to  dry  return,  the  pressure  drops  to  two  different  points 
would  not  necessarily  be  the  same  and  the  system  would  short  circuit. 

10  •  Why  cannot  graduated  valves  be  used  on  a  one-pipe  system? 

Partial  opening  of  valves  would  restrict  flow  to  such  an  extent  that  the  radiator  could  not 
drain  properly  and  would  fill  with  water. 

11  •  What  advantage  is  there  to  an  air  valve  with  a  check  to  prevent  the  re- 
entrance  of  expelled  air? 

A  system  equipped  with  such  valves  builds  up  a  vacuum  and  holds  the  heat  longer. 
With  proper  controls  on  the  boiler,  lower  radiator  temperatures  can  be  maintained  in 
mild  weather,  giving  better  plant  efficiency. 

12  •  With  a  one-pipe  steam  heating  plant  designed  for  a  total  pressure  drop  of 
J4  Ib  with  a  check  valve  on  the  return,  how  high  must  the  lowest  part  of  the 
steam  main  be  aboye  the  boiler  water  line? 

Water  line  difference  (Ji  X  28)  7  in. 

Flow  head  required  3  in. 

Friction  head  of  check  valve  4  in. 

Factor  of  safety  6  in. 


Total  required  20  in, 

13  •  What  are  the  essentials  of  a  two-pipe  closed  vapor  system? 

Packless  graduated  valves  on  radiators ;  thermostatic  return  traps  on  returns  and  drips ; 
an  automatic  return  trap  to  prevent  water  from  backing  out  of  the  boiler. 

14  •  Why  must  the  automatic  return  trap  on  two-pipe  vapor  systems  be  about 
18  in.  above  the  boiler  water  line? 

That  height  is  necessary  to  overcome  water  line  difference  owing  to  pressure  drop  and 
friction  in  pipe  and  fittings. 

15  •  What  is  the  difference  between  the  systems  illustrated  in  Fig.  10  and 
Fig.  13? 

The  risers  and  the  air  eliminator  in  Fig,  13  are  vented  to  atmosphere. 

525 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

16  •  What  is  the  difference  between  a  vacuum  pump  and  a  condensation  return 
pump? 

The  vacuum  pump  produces  and  maintains  a  vacuum  in  the  return  lines  whereas  the 
condensation  return  pump  returns  the  condensation  back  to  the  boiler.  The  relation  of 
the  boiler  to  the  heating  units  is  such  that  the  condensate  will  not  return  by  gravity. 

17  •  What  is  the  function  of  a  trap? 

The  usual  function  is  to  allow  the  passage  of  condensate  and  air  and  to  prevent  the 
passage  of  steam. 

18  •  Under  what  conditions  is  it  advisable  to  use  a  combination  float  and 
thermostatic  trap? 

Where  unusual  capacities  are  required,  as  on  large  mains  or  blast  coils. 

19  •  Why  should  the  discharge  from  high  pressure  traps  not  go  directly  into  a 
vacuum  return  main? 

Because  of  its  higher  temperature,  the  high  pressure  condensate  would  immediately 
flash  into  steam  in  the  vacuum  main  and  cause  difficulty  with  the  vacuum  pump. 

20  •  What  is  the  function  of  the  automatic  return  trap? 

To  insure  the  return  of  condensate  to  the  boiler  when  the  operating  condition  is  such  that 
the  boiler  pressure  exceeds  the  static  head  on  the  returns. 


526 


Chapter  32 

PIPING  FOR  STEAM  HEATING 

SYSTEMS 

Flow  of  Steam  in  Pipes,  Pipe  Sizes,  Tables  Jor  Pipe  Sizing,  Sizing 
One-Pipe  Gravity  Air  Vent  Systems,  Two-Pipe  Gravity  Air  Vent 
Systems,  Two-Pipe  Vapor  Systems,  Atmospheric  Systems, 
Vacuum  Systems,  Sub -Atmospheric  Systems,  Orifice  Systems, 
High  Pressure  Steam,  Expansion  in  Steam  and  Return  Lines, 
Piping  Connections  and  Details,  Boiler  Connections,  Hartford 
Return  Connection 

THE  design  of  a  steam  heating  system  should  be  considered  under  four 
headings,  namely,  (1)  the  details  of  the  heating  units,  (2)  the  arrange- 
ment of  the  general  piping  scheme,  (3)  the  details  of  connections,  and  (4) 
the  sizing  of  the  lines.    Items  1  and  2  are  covered  in  Chapters  30  and  31 ,. 
respectively,  while  this  chapter  considers  the  two  latter  items. 

The  functions  of  piping  are  to  supply  the  heating  units  with  steam  and 
to  remove  the  condensation.  In  some  systems  both  the  air  and  con- 
densation are  removed  from  the  heating  units  by  the  return  piping.  To 
accomplish  this  effectively,  the  distribution  of  the  steam  should  be 
efficient  and  equitable,  without  noise,  and  the  returns  should  be  as  short 
as  possible.  When  air  is  handled  its  escape  should  be  facilitated  to  the 
utmost  since  an  air-bound  system  will  not  heat  properly.  Condensation 
takes  place  in  a  steam  system  not  only  in  the  heating  units,  but  through- 
out the  piping  system  as  well,  and  the  returns  also  condense  any  steam  or 
vapor  that  may  be  contained.  At  the  same  time  part  of  the  condensation 
may  flash  back  into  steam  when  the  vacuum  or  pressure  in  the  return  is- 
considerably  below  the  steam  pressure. 

It  is  essential  that  steam  piping  systems  not  only  distribute  steam  at 
full  load  but  also  at  partial  loads,  as  the  average  winter  demand  is  less 
than  half  of  the  demand  in  most  severe  outside  temperatures.  Further- 
more, in  heating  up  rapidly  the  load  on  the  steam  main  may  exceed  the 
maximum  operating  load  even  in  extreme  weather,  due  to  the  necessity 
of  raising  the  temperature  of  the  metal  in  the  system  to  the  steam  tem- 
perature. This  may  require  more  heat  than  would  be  emitted  from  the 
system  itself  after  it  once  is  thoroughly  heated. 

STEAM  FLOW 

The  rate  of  flow  of  dry  steam  or  steam  with  a  small  amount  of  water 
flowing  in  the  same  direction  is  in  accordance  with  the  general  laws  of  gas 
flow  and  is  a  function  of  the  length  and  diameter  of  the  pipe,  the  density 
of  the  steam,  and  the  pressure  drop  through  the  pipe.  This  relationship 
has  been  established  by  Babcock  in  the  following  formula: 

527 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


P  =  0.0000000367  (  1  4-  ~p\  -^  (1) 

or 

W  =  5220  , 


PDd* 

77+^) 


where 

P  —  loss  in  pressure,  pounds  per  square  inch. 

d  —  inside  diameter  of  pipe,  inches. 

L  ~  length  of  pipe,  feet. 

D  —  weight  of  1  cu  ft  of  steam. 

W  ~  weight  of  steam  flowing  per  hour,  pounds. 

Example  1.    How  much  steam  will  flow  per  hour  through  100  ft  of  2-in.  pipe  if  the 
initial  pressure  is  1.3  Ib  per  square  inch  and  the  pressure  drop  is  1  02? 

Solution.    P  «  ~  =  0.0625  Ib;  d  =  2.067  in.  (Table   1,  Chapter  34);  L  =  100  ft; 

ID 
D  =  0.04038  Ib  (Table  7,  Chapter  1).    Substituting  these  values  in  Formula  2: 


V 0 


J  0.0625  X  0.04038  X  2.0675 
7  3^~T  =  97.2  Ib  per  hour. 


Formula  2  does  not  allow  for  entrained  water  in  low-pressure  steam, 
condensation  in  pipe,  and  roughness  in  commercial  pipe  as  found  in 
practice. 

The  latent  heat  of  steam  (/zfg)  at  atmospheric  pressure  (Table  7, 
Chapter  1)  is  970.2  Btu  per  pound.  Inasmuch  as  the  heat  emission  of  an 
equivalent  square  foot  of  heating  surface  (radiation)  is  240  Btu,  1  Ib  of 

970  2 

steam  at  this  pressure  will  supply   ~ .  *>••  or  4.04  sq  ft  of  equivalent  heating 

surface.  This  figure  is  usually  taken  as  4  even.  In  Example  1,  the  weight 
of  steam  flowing  per  hour  would  therefore  supply  4  X  97.2  or  388.8  sq  ft 
of  equivalent  heating  surface. 

PIPE  SIZES 

The  determination  of  pipe  sizes  for  steam  heating  depends  on  the 
following  principal  factors : 

1 .  The  initial  pressure  and  the  total  pressure  drop  which  may  be  allowed  between  the 
source  of  supply  and  the  end  of  the  return  system. 

2.  The  maximum  velocity  of  steam  allowable  for  quiet  and  dependable  operation  of 
the  system. 

3.  The  equivalent  length  of  the  run  from  the  boiler  or  source  of  steam  supply  to  the 
farthest  heating  unit. 

4.  Unusual  conditions  in  the  building  to  be  heated. 

Initial  Pressure  and  Pressure  Drop 

Theoretically  there  are  several  factors  to  be  considered,  such  as  initial 
pressure  and  pressure  required  at  the  end  of  the  line,  but  it  is  most  im- 
portant that  (1)  the  total  pressure  drop  does  not  exceed  the  initial  pressure 
of  the  system ;  (2)  the  pressure  drop  is  not  so  great  as  to  cause  excessive 
velocities;  (3)  there  is  a  constant  initial  pressure,  except  on  systems 

528 


CHAPTER  32 — PIPING  FOR  STEAM  HEATING  SYSTEMS 


TABLE  1.     MAXIMUM  ALLOWABLE  CAPACITIES  OF  UP-FEED  RISERS  FOR  ONE-PIPE 

Low  PRESSURE  STEAM 

Based  on  A.  S.  H.  V.  E*  Research  Laboratory  Tests 


PIPE  SIZE 
INCHRS 

VELOCITY 

FEET  PER  SECOND 

PRESSURE  DROP 
OUNCES 
PER  100  FT 

CAPACITY 

SgFt 
Radiation 

Btu  per  Hour 

Lb 
Steam  per  Hour 

A 

B 

C 

D 

E 

F 

1 

14.1 

0.68 

45 

10,961 

11.3 

1J€ 

17.6 

0.66 

98 

23,765 

24.5 

i« 

20.0 

0.66 

152 

36,860 

38.0 

2 

23.0 

0.57 

2S8 

69,840 

72.0 

2H 

26.0 

0.54 

464 

112,520 

116.0 

3 

29.0 

0.48 

799 

193,600 

199.8 

3H 

31.0 

0.44 

1144 

277,000 

286.0 

4 

32.0 

0.39 

1520 

368,000 

380.0 

INSTRUCTIONS  FOR  USING  TABLE  1 

1.  Capacities  given  in  Table  1  should  never  be  exceeded  on  one-pipe  risers. 

2.  Capacities  are  based  on  J<-lb  condensation  per  square  foot  equivalent  radiation  and  actual  diameter 
of  standard  pipe. 

3.  All  pipe  should  be  well  reamed  and  free  from  constrictions.    Fittings  should  be  up  to  size.     (See 
Tables  4  and  5). 

specially  designed  for  varying  initial  pressures,  such  as  the  sub-atmos- 
pheric, the  orifice,  and  the  vapor  systems  which  normally  operate  under 
partial  vacuums;  (4)  there  is  sufficient  difference  in  level,  for  gravity 
return  systems,  between  the  lowest  point  on  the  steam  main,  the  heating 
units,  and  the  dry  return,  when  considered  in  relation  to  the  boiler  water 
line. 

All  systems  should  be  designed  for  a  low  initial  pressure  and  a  reason- 
ably small  pressure  drop  for  two  reasons :  first,  the  present  tendency  in 
steam  heating  unmistakably  points  toward  a  constant  lowering  of  pres- 
sures even  to  those  below  atmospheric;  second,  a  system  designed  in  this 
manner  will  operate  under  higher  pressures  without  difficulty.  When  a 
system  designed  for  a  relatively  high  initial  pressure  and  a  relatively  high 
pressure  drop  is  operated  at  a  lower  pressure,  it  is  likely  to  be*  noisy  and 
have  poor  circulation. 

The  total  pressure  drop  should  never  exceed  one-half  of  the  initial 
pressure  when  condensate  is  flowing  in  the  same  direction  as  the  steam. 
Where  the  condensate  must  flow  counter  to  the  steam,  the  governing 
factor  is  the  velocity  permissible  without  interfering  with  the  condensate 
flow.  Laboratory  experiments  limit  this  to  the  capacities  given  in 
Tables  1  and  2  for  vertical  risers  and  in  Table  3  for  horizontal  pipes  at 
varying  grades. 

Maximum  Velocity  and  Reaming 

The  capacity  of  a  steam  pipe  in  any  part  of  a  steam  system  depends 
upon  the  quantity  of  condensation  present,  the  directipa  in  which  the 
condensate  is  flowing,  and  the  pressure  drop  in  the  pipe.  Where  the 

£29 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


TABLE  2.    MAXIMUM  ALLOWABLE  CAPACITIES  OF  UP-FEED  RISERS  FOR  Two- PIPE 

Low  PRESSURE  STEAM 

Based  on  A.  S.  H.  V.E.  Research  Laboratory  Tests 


PIPE  SIZE 
INCHES 

VELOCITY 
FEET  PER  SECOND 

PRESSURE  DROP 
OUNCES 
PER  100  FT 

CAPACITY 

SqFt 
Radiation 

Btu  per  Hour 

Lb 
Steam  per  Hour 

A 

B 

C 

D 

E 

F 

% 

20 



40 

9550 

10.0 

1 

23 

1.78 

74 

17,900 

18.45 

IJi 

27 

1.57 

151 

36,500 

37.65 

1J* 

30 

1.48 

228 

55,200 

57.0 

2 

35 

1.33      • 

438 

106,100 

109.5 

m 

38 

1.16 

678 

164,100 

169.4 

3 

41 

0.95 

1129 

273,500 

282.2 

3^ 

42 

0.81 

1548 

375,500 

387.0 

4 

43 

0.71 

2042 

495,000 

510.5 

INSTRUCTIONS  FOR  USING  TABLE  2 

1  .  The  capacities  given  in  this  table  should  never  be  exceeded  on  two-pipe  risers. 

2.  Capacities  are  based  on  J^-lb  condensation  per  square  foot  equivalent  radiation  and  actual  diameter 


of  standard  pipe. 

3.  AH  pipe  should  be  well  reamed  and  free  from  constrictions. 
Tables  4  and  5.) 


Fittings  should  be  up  to  size.     (See 


quantity  of  condensate  is  limited  and  is  flowing  in  the  same  direction  as 
the  steam,  only  the  pressure  drop  need  be  considered.  When  the  con- 
densate must  flow  against  the  steam,  even  in  limited  quantity,  the  ve- 
locity of  the  steam  must  not  exceed  limits  above  which  the  disturbance 
between  the  steam  and  the  counter-flowing  water  may  produce  object- 
ionable sounds,  such  as  water  hammer,  or  may  result  in  the  retention  of 
water  in  certain  parts  of  the  system  until  the  steam  flow  is  reduced 
sufficiently  to  permit  the  water  to  pass.  The  velocity  at  which  such 
disturbances  take  place  is  a  function  of  (1)  the  pipe  size,  whether  the  pipe 
runs  horizontally  or  vertically,  (2)  the  pitch  of  the  pipe  if  it  runs  hori- 
zontally, and  (3)  the  quantity  of  condensate  flowing  against  the  steam. 
Two  factors  of  uncertainty  always  exist  in  determining  the  capacity  of 
any  steam  pipe.  The  first  is  variation  in  manufacture,  which  apparently 
cannot  be  avoided  and  which  caused  an  actual  difference  of  20  per  cent  in 
the  capacity  of  a  1-in.  pipe  in  experiments  carried  on  at  the  A.S.H.V.E. 
Research  Laboratory  (Table  4).  The  second  is  the  reaming  of  the  ends  of 
the  pipe  after  cutting,  which,  experiments  indicate,  might  reduce  the 
capacity  of  a  1-in.  pipe  as  much  as  28.7  per  cent  (Table  5).  All  of  the 
capacity  tables  given  in  this  chapter  include  a  factor  of  safety.  However, 
the  pipe  on  which  Table  4  is  based  showed  no  particular  defects  or  con- 
strictions on  the  inside,  and  the  factor  of  safety  referred  to  does  not  cover 
abnormal  defects  or  constrictions  nor  does  it  cover  pipe  not  properly 
reamed. 

530 


CHAPTER  32 — PIPING  FOR  STEAM  HEATING  SYSTEMS 


TABLE  3.     COMPARATIVE  CAPACITY  OF  STEAM  LINES  AT  VARIOUS  PITCHES'* 
Pitch  of  Pipe  in  Inches  per  10  Ft 


PITCH  OP 

1 

PIPE 

1A    IN. 

1A  IN. 

1  IN. 

IHm. 

2  IN. 

3  IN. 

4  IN. 

5  IN. 

Pipe 

SqFt 
Rad. 

3 

SqFt 
Rad. 

? 

SqFt 
Rad. 

"3 

SqFt 
Rad. 

*« 

SqFt 
Rad. 

13 

SqFt 
Rad. 

"3 

SqFt 
Rad. 

"3 

SqFt 
Rad. 

•a 

Size 

Based 

K 

Based 

M 

Based 

^ 

Based 

S* 

Based 

V 

Based 

'>. 

Based 

^ 

Based 

. 

Inches 

on  240 
Btu 

1 

on  240 
Btu 

1 

on  240 
Btu 

1 

on  240 
Btu 

on  240 
Btu 

§ 

on  240 
Btu 

s 

on  240 
Btu 

C3 

on  240 
Btu 

1 

H 

25.0 

12 

30.3 

14 

37.3 

18 

40.4 

19 

42.5 

20 

46.1 

21 

47.5 

22 

49.3 

23 

i 

45.8 

12 

52.6 

15 

63.0 

17 

70.0 

20 

75.2 

22 

83.0 

23 

87.9 

25 

90.2 

26 

ij^ 

104.9 

18 

117.2 

20 

133.0 

23 

144.5 

25 

154.0 

27 

165.0 

28 

172.6 

29 

178.2 

31 

ij^ 

142.6 

18 

159.0 

21 

181.0 

23 

196.5 

25 

209.3 

27 

224.0 

28 

234.8 

30 

242.6 

31 

2 

236.0 

19 

263.5 

20 

299.5 

23 

325.5 

25 

346.5 

27 

371.5 

28 

388.4 

29 

401,1 

30 

aData  from  A.S.H.V.E.  Research  Laboratory. 

Equivalent  Length  of  Run 

All  tables  for  the  flow  of  steam  in  pipes,  based  on  pressure  drop,  must 
allow  for  the  friction  offered  by  the  pipe  as  well  as  for  the  additional 
resistance  of  the  fittings  and  valves.  These  resistances  generally  are 
stated  in  terms  of  straight  pipe;  in  other  words,  a  certain  fitting  will 
produce  a  drop  in  pressure  equivalent  to  so  many  feet  of  straight  run  of 
the  same  size  of  pipe.  Table  6  gives  the  number  of  feet  of  straight  pipe 
usually  allowed  for  the  more  common  types  of  fittings  and  valves.  In  all 
pipe  sizing  tables  in  this  chapter  the  length  of  run  refers  to  the  equivalent 
length  of  run  as  distinguished  from  the  actual  length  of  pipe  in  feet.  The 
length  of  run  is  not  usually  known  at  the  outset;  hence  it  is  necessary  to 
assume  some  pipe  size  at  the  start.  Such  an  assumption  frequently  is 
considerably  in  error  and  a  more  common  and  practical  method  is  to 
assume  the  length  of  run  and  to  check  this  assumption  after  the  pipes  are 
sized.  For  this  purpose  the  length  of  run  usually  is  taken  as  double  the 
actual  length  of  pipe. 

TABLE  4.    PER  CENT  DIFFERENCE  IN  CAPACITY  FOR  CARRYING  STEAM  AND  CONDENSATE 
DUE  TO  VARIATION  OF  PIPE  SIZE  AND  SMOOTHNESS**- 


MAXIMUM  CONDENSATION,  LB  PER  HOUR 


Size  of  pipe  

5*  In. 

lln. 

Itfln. 

iMIn. 

Minimum.  

14.00 

24.89 

45.42 

70.50 

Maximum  

15.20 

30.08 

52.08 

82.00 

Per  cent  variation  

8.6 

20.8 

14.7 

16.3 

aData  from  AMERICAN  SOCIETY  OF  HEATING  AND  VENTILATING  ENGINEERS  Research  Laboratory. 

TABLE  5.    EFFECT  OF  REAMING  ENTRANCE  TO  ONE-INCH  ONE-PIPE  RISERS* 


MAXIMUM  CAPACITY 
OF  RISER 

PER  CENT 
DECREASE 

Reamed  entrances                   ,                   ... 

24.7  Ib  per  hour 

0.0 

Rounded  entrances          

23.9  Ib  per  hour 

3.2 

Squared  entrances  

22.2  Ib  per  hour 

10.1 

Three  wheel  cutter     .        

19.2  Ib  per  hour 

22.2 

Single  wheel  cutter 

17,6  Ib  per  hour 

28.7 

aData  from  AMERICAN  SOCIETY  OF  HEATING  AND  VENTILATING  ENGINEERS  Research  Laboratory. 

531 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

TABLE  6.    LENGTH  IN  FEET  OF  PI?E  TO  BE  ADDED  TO  ACTUAL  LENGTH  OF  RUN — 
OWING  TO  FITTINGS — TO  OBTAIN  EQUIVALENT  LENGTH 


SIZE  OF  PIPE 
INCHES 

ST'D.  ELBOW 

SIDE  OUTLET 

TEE 

GATE  VALVE 

GLOBE  VALVE 

ANGLE  VALVE 

Length  in  Feet  to  be  Added  to  Run 

2 

5 

16 

2 

18 

9 

VA 

7 

20 

3 

25 

12 

3 

10 

26 

3 

33 

16 

3^ 

12 

31 

4 

39 

19 

4 

14 

35 

5 

45 

22 

5 

18 

44 

7 

57 

28 

6 

22 

50 

9 

70 

32 

7 

26 

55 

10 

82 

37 

8 

31 

63 

12 

94 

42 

9 

35 

69 

13 

105 

47 

10 

39 

76 

15 

118 

52 

12 

47 

90 

18 

140 

63 

14 

53 

105 

20 

160 

72 

Example  of  length  in 
feet  of  pipe  to  be  added 
to  actual  length  of  run. 


LM5TH.    - 

• 

\    !  EQUIVALENT  LEH6TH '-  19$ -0* 


TABLES  FOR  PIPE  S1ZINQ1 

Factors  determining  the  size  of  a  steam  pipe  and  its  allowable  limit  of 
capacity  are  as  follows: 

1.  Pipe  condensate  flowing  with  steam. 

2.  Pipe  condensate  flowing  against  steam. 

3.  Pipe  and  radiator  condensate  flowing  with  steam. 

4.  Pipe  and  radiator  condensate  flowing  against  steam. 

It  is  apparent  that  (3)  and  (4)  are  practically  limited  to  one-pipe 
systems  while  (1)  and  (2)  cover  all  other  systems. 

Tables  7  and  8,  worked  out  for  determining  pipe  sizes,  have  their  col- 
umns lettered  continuously,  Columns  A  through  L  being  in  Table  7,  and 
M  through  EE  in  Table  8.  In  the  following  text,  reference  made  to 
columns  will  be  by  letter.  The  tables  are  based  on  the  actual  inside 
diameters  of  the  pipe  and  the  condensation  of  ^  Ib  (4  oz)  of  steam  per 
square  foot  of  equivalent  direct  radiation2  (abbreviated  EDR)  per  hour. 
The  drops  indicated  are  drops  in  pressure  per  100  ft  of  equivalent  length 
of  run.  The  pipe  is  assumed  to  be  well  reamed  without  unusual  or  notice- 
able defects. 


iPipe  size  tables  in  this  chapter  have  been  compiled  in  simplified  and  condensed  form  for  the  convenience 
of  the  user:  at  the  same  time  all  of  the  information  contained  in  previous  editions  of  THE  GUIDE  has  been 
retained.  Values  of  pressure  drops,  formerly  expressed  in  ounces,  are  now  expressed  in  fractions  of  a  pound. 

aAa  steam  system  design  has  materially  changed  in  recent  years  so  that  240  Btu  no  longer  expresses  the 
heat  of  condensation  from  a  square  foot  of  radiator  surface  per  hour,  and  as  present  day  heating  units  have 
different  characteristics  from  older  forms  of  radiation,  it  is  the  purpose  of  THE  GUIDE  to  gradually  eliminate 
the  empirical  expression  square  foot  of  equivalent  direct  radiation,  EDR,  and  to  substitute  a  logical  unit  based 
on  the  Btu,  The  new  terms  to  express  the  equivalent  of  1000  Btu  (Mb),  and  1000  Btu  per  hour  (Mbh), 
have  been  approved  by  the  A.S.H.V.E. 


CHAPTER  32 — PIPING  FOR  STEAM  HEATING  SYSTEMS 


TABLE  7.    STEAM  PIPE  CAPACITIES 

Capacity  Expressed  in  Square  Feet  of  Equivalent  Direct  Radiation 
(Reference  to  this  table  will  be  by  column  letter  A.  through  L) 

This  table  is  based  on  pipe  size  data  developed  through  the  research  investiga- 
tions of  the  AMERICAN  SOCIETY  OF  HEATING  AND  VENTILATING  ENGINEERS. 


CAPACITIES  OF  STEAM  MAINS  AND    RISERS 


PIPE 
SIZE 
IN. 

DIRECTION  OF  CONDENSATION  FLOW  IN  PIPE  LINE 

With  the  Steam  in  One-Pipe  and  Two-Pipe  Systems 

Against  the  Steam 
Two-Pipe  Only 

Supply 
Risers 
Up- 
Feed 

Radiator 
Valves 
and 
Vertical 
Con- 
nections 

Radiator 
and 
Riser 
Run- 
outs 

Vsa  Ib 
or 
HOz 
Drop 

1/24  Ib 

or 
^Oz 
Drop 

Vie  Ib 
or 
1  Oz 
Drop 

Hlo 

or 
20z 
Drop 

Mlb 
or 
40z 
Drop 

>Ub 
or 
8  Oz 
Drop 

Vertical 

Hori- 
zontal 

A 

B 

C 

D 

E 

P 

G 

#a 

fc 

/b 

K 

Lc 

H 
1 

IX 

iy* 

2 

m 

3^ 
4 
5 
6 
8 
10 
12 
16 

30 
56 

122 
190 
386 
635 
1,163 
1,737 
2,457 
4,546 
7,462 
15,533 
28,345 
45,492 
84,849 

~jg 
173 
269 
546 
898 
1,645 
2,457 
3,475 
6,429 
10,553 
21,967 
40,085 
64,336 
121,012 

30 
56 
122 
190 
386 
635 
1,129 
1,548 
2,042 

"""26 
58 
95 
195 
395 
700 
1,150 
1,700 
3,150 

25 
45 
98 
152 
288 
464 
799 
1,144 
1,520 

39 

87 
134 
273 
449 
822 
1,228 
1,738 
3,214 
5,276 
10,983 
20,043 
32,168 
60,506 

46 
100 
155 
315 
518 
948 
1,419 
2,011 
3,712 
6,094 
12,682 
23,144 
37,145 
69,671 

111 

245 
380 
771 
1,270 
2,326 
3,474 
4,914 
9,092 
14,924 
31,066 
56,689 
90,985 
169,698 

157 
346 
538 
1,091 
1,797 
3,289 
4,913 
6,950 
12,858 
21,105 
43,934 
80,171 
128,672 
242,024 

20 
55 
81 
165 

20 
55 
81 
165 
260 
475 
745 
1,110 
2,180 

All  Horizontal  Mains  and  Down-Feed  Risers 

3*1 

Feed 

Risers 

Mains 
and  Un- 
dripped 
Run- 
outs 

Up- 
Fe£ 
Risers 

Radiator 
Con- 
nections 

Run- 
outs 
Not 
Dripped 

SPECIAL  CAPACITIES  FOR 


te. — All  drops  shown  are  in  pounds  per  100  ft  of  equivalent  run — based  on  pipe  properly  reamed. 
aDo  not  use  Column  H  for  drops  of  1/24  or  1/32  Ib;  substitute  Column  C  or  Column  B  as  required, 
bDo  not  use  Column  J  for  drop  of  1/32  Ib  except  on  sizes  3  in.  and  over;  below  3  in.  substitute  Column  B. 
cOn  radiator  runouts  over  8  ft  long  increase  one  pipe  size  over  that  shown  in  Table  7. 
^      ~s*^+  I  AMERICAN  SOCIETY  OF  HEATING  AND  VENTILATING  ENGINEERS       )   Not  to  be  Reprinted  With- 
Copyright  |  Heaiin{Jt  Piping  and  Air  Cmdtiiming  Cmtractns  Naticmd  Astutoion  $  out  Special  Permission 

Table  7  may  be  used  for  sizing  piping  for  steam  heating  systems  by 
determining  the  allowable  or  desired  pressure  drop  per  100  equivalent 
feet  of  run  and  reading  from  the  column  for  that  particular  pressure  drop. 
This  applies  to  all  steam  mains  on  both  one-pipe  and  two-pipe  systems, 
vapor  systems,  and  vacuum  systems.  Columns  B  to  G,  inclusive,  are  used 
where  the  steam  and  condensation  flow  in  the  same  direction,  while 
Columns  H  and  /  are  for  cases  where  the  steam  and  condensation  flow  in 
opposite  directions,  as  in  risers  and  runouts  that  are  not  dripped.  Columns 
J,  K,  and  L  are  for  one-pipe  systems  and  cover  riser,  radiator  valve,  and 
vertical  connection  sizes,  and  radiator  and  runout  sizes,  all  of  which  are 
based  on  the  critical  velocities  of  the  steam  to  permit  the  counter  flow  of 
condensation  without  noise. 

Sizing  of  return  piping  may  be  done  with  the  aid  of  Table  8  where  pipe 
capacities  for  wet,  dry,  and  vacuum  return  lines  are  shown  for  the  pressure 
drops  per  100  ft  corresponding  to  the  drops  in  Table  7.  It  is  customary  to 
use  the  same  pressure  drop  on  both  the  steam  and  return  sides  of  a  system. 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


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a 


CHAPTER  32 — PIPING  FOR  STEAM  HEATING  SYSTEMS 


Example  2.  What  pressure  drop  should  be  used  for  the  steam  piping  of  a  system  if 
the  measured  length  of  the  longest  run  is  500  ft  and  the  initial  pressure  is  not  to  be  over 
2-lb  gage? 

Solution.  It  will  be  assumed,  if  the  measured  lengtn  of  the  longest  run  is  500  ft,  that 
when  the  allowance  for  fittings  is  added  the  equivalent  length  of  run  will  not  exceed 
1,000  ft.  Then,  with  the  pressure  drop  not  over  one  half  of  the  initial  pressure,  the  drop 
could  be  1  Ib  or  less.  With  a  pressure  drop  of  1  Ib  and  a  length  of  run  of  1,000  ft,  the 
drop  per  100  ft  would  be  J-f  0  Ib,  while  if  the  total  drop  were  1A  Ib,  the  drop  per  100  ft 
would  be  3^o  Ib.  In  the  first  instance  the  pipe  could  be  sized  according  to  Column  D  for 
He  Ib  per  100  ft,  and  in  the  second  case,  the  pipe  could  be  sized  according  to  Column  C 
for  J^4  Ib.  On  completion  of  the  sizing,  the  drop  could  be  checked  by  taking  the  longest 
line  and  actually  calculating  the  equivalent  length  of  run  from  the  pipe  sizes  determined. 
If  the  calculated  drop  is  less  than  that  assumed,  the  pipe  size  is  all  right;  if  it  is  more,  it  is 
probable  that  there  are  an  unusual  number  of  fittings  involved,  and  either  the  lines  must 
be  straightened  or  the  column  for  the  next  lower  drop  must  be  used  and  the  lines  resized. 
Ordinarily  resizing  will  be  unnecessary. 

.      ONE-PIPE  GRAVITY  AIR-VENT  SYSTEMS 

One-pipe  gravity  air-vent  systems  in  which  the  equivalent  length  of  run 
does  not  exceed  200  ft  should  be  sized  as  follows : 

1.  For  the  steam  main  and  dripped  runouts  to  risers  where  the  steam  and  condensate 
flow  in  the  same  direction,  use  Jie-lb  drop  (Column  D}. 

2.  Where  the  riser  runouts  are  not  dripped  and  the  steam  and  condensation  flow  in 
opposite  directions,  and  also  in  the  radiator  runouts  where  the  same  condition  occurs,  use 
Column  L. 

3.  For  up-feed  steam  risers  carrying  condensation  back  from  theradiators,  use  Column  /. 

4.  For  dawn-feed  systems  the  main  risers  of  which  do  not  carry  any  radiator   con- 
densation, use  Column  H. 

5.  For  the  radiator  valve  size  and  the  st^^b  connection,  use  Column  X. 

6.  For  the  dry  return  main,  use  Column  U. 

7.  For  the  ivet  return  main  use  Column  T. 

On  systems  exceeding  an  equivalent  length  of  200  ft,  it  is  suggested  that 
the  total  drop  be  not  over  ^  Ib.  The  return  piping  sizes  should  correspond 
with  the  drop  used  on  the  steam  side  of  the  system.  Thus,  where  ^44b 
drop  is  being  used,  the  steam  main  and  dripped  runouts  would  be  sized  from 
Column  C;  radiator  runouts  and  undripped  riser  runouts  from  Column  L; 
up-feed  risers  from  Column  J;  the  main  riser  on  a  down-feed  system  from 
Column  C  (it  will  be  noted  that  if  Column  H  is  used  the  drop  would 
•exceed  the  limit  of  J^4  Ib) ;  the  dry  return  from  Column  R;  and  the  wet 
return  from  Column  Q. 

With  a  J^2-lb  drop  the  sizing  would  be  the  same  as  for  H*  Ib  except  that 
the  steam  main  and  dripped  runouts  would  be  sized  from  Column  B,  the 
main  riser  on  a  down-feed  system  from  Column  B,  the  dry  return  from 
Column  0,  and  the  wet  return  from  Column  N. 

Example  8,  Size  the  one-pipe  gravity  steam  system  shown  in  Fig.  1  assuming  that 
this  is  all  there  is  to  the  system  or  that  the  riser  and  run  shown  involve  the  longest  run 
on  the  system. 

Solution.  The  total  length  of  run  actually  shown  is  215  ft.  If  the  equivalent  length 
of  run  is  taken  at  double  this,  it  will  amount  to  430  ft,  and  with  a  total  drop  of  J4  Ib 
the  drop  per  100  ft  will  be  slightly  less  than  He  Ib.  It  would  be  well  in  this  case  to  use 
3^24  Ib,  and  this  would  result  in  the  theoretical  sizes  indicated  in  Table  9,  These  theo- 

535 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


TABLE  9. 


PIPE  SIZES  FOR  ONE-PIPE  UP-FEED  SYSTEM  SHOWN 
IN  FIG.  1 


PART  OF  SYSTEM 

SECTION 
OP  PIPE 

RADIATION 
SUPPLIED 
(Sq  FT) 

THEORETICAL 
PIPE  SIZE 
(INCHES) 

PRACTICAL 
PIPE  SIZE 
(INCHES) 

Branches  to  radiators- 
Branches  to  radiators.. 
Riser  

a  to  & 

100 
50 
200 

2 

llA 

2 

2 

Ui 
2 

Riser 

b  to  c 

300 

2^ 

21A 

Riser  ..            

c  to  d 

400 

2H 

2M 

Riser 

d  to  e 

500 

3 

3 

Riser  

etof 

600 

3 

3 

Branch  to  riser 

/to  £ 

600 

3^2 

3M 

Supply  main  

J    ^  5 

g  to  h 

600 

3 

3 

Branch  to  supply  main 
Dry  return  main  

htoj 
ftok 

600 
600 

2H 

11A 

3 
2 

Wet  return  main  

ktom 

600 

1 

2 

Wet  return  main  
Wet  return  main  

m  to  n 
n  to  p 

600 
'600 

1 
i 

2 

2 

FIG.  1.      RISER,  SUPPLY 

MAIN  AND  RETURN  MAIN 

OF  ONE- PIPE  SYSTEM 


retical  sizes,  however,  should  be  modified  by  not  using  a  wet  return  less  than  2  in.  while 
the  main  supply,  g-h,  if  from  the  uptake  of  a  boiler,  should  be  made  the  full  size  of  the 
main,  or  3  in.  Also  the  portion  of  the  main  k-m  should  be  made  2  in.  if  the  wet  return 
is  made  2  in. 

Notes  on  Gravity  One- Pipe  Air- Vent  Systems 

1.  Radiator  runouts  over  8  ft  long  should  be  increased  one  pipe  size. 

2.  Pitch  of  mains  should  be  not  less  than  %  in.  in  10  ft. 

3.  Pitch  of  horizontal  runouts  to  risers  and  radiators  should  not  be  less  than  J-£  in. 
in  10  ft. 

4.  In  general,  it  is  not  desirable  to  have  a  main  less  than  2  in.    The  diameter  of  the 
far  end  of  the  supply  main  should  be  not  less  than  half  its  diameter  at  its  largest  part. 

5.  Supply  mains,  branches  to  risers,  or  risers,  should  be  dripped  where  necessary. 

TWO-PIPE  GRAVITY  AIR-VENT  SYSTEMS 

The  method  employed  in  determining  pipe  sizes  for  two-pipe  gravity 
air-vent  systems  is  similar  to  that  described  for  one-pipe  systems  except 
that  the  steam  mains  never  carry  radiator  condensation.  The  drop 
allowable  per  100  ft  of  equivalent  run  is  obtained  by  taking  the  equiva- 
lent length  to  the  farthest  radiator  as  double  the  actual  distance,  and 
then  dividing  the  allowable  or  desired  total  drop  by  the  number  of 
hundreds  of  feet  in  the  equivalent  length.  Thus  in  a  system  measuring 
400  ft  from  the  boiler  to  the  farthest  radiator,  the  approximate  equivalent 
length  of  run  would  be  800  ft-  With  a  total  drop  of  J^  Ib  the  drop  per 

100  ft  would  be  ~  or  ^6  Ib;  therefore,  Column  D  would  be  used  for  all 

o 

steam  mains  where  the  condensation  ,and  steam  flow  in  the  same  direc- 
tion. If  a  total  drop  of  %  Ib  is  desired,  the  drop  per  100  ft  would  be  ^2  Ib 

536 


CHAPTER  32 — PIPING  FOR  STEAM  HEATING  SYSTEMS 


and  Column  JB  would  be  used.  If  the  total  drop  were  to  be  1  Ib,  the  drop 
per  100  ft  would  be  ^  Ib  and  Column  E  would  be  used. 

For  mains  and  riser  runouts  that  are  not  dripped,  and  for  radiator 
runouts  where  in  all  three  cases  the  condensation  and  steam  flow  in 
opposite  directions,  Column  I  should  be  used,  while  for  the  steam  risers 
Column  H  should  be  used  unless  the  drop  per  100  ft  is  $$4  Ib  or  ^  Ib, 
when  Columns  B  or  C  should  be  substituted  so  as  not  to  exceed  the  drop 
permitted. 

On  an  overhead  down-feed  system  the  main  steam  riser  should  be 
sized  by  reference  to  Column  H,  but  the  down-feed  steam  risers  sup- 
plying the  radiators  should  be  sized  by  the  appropriate  Columns  B  through 
G,  since  the  condensation  flows  downward  with  the  steam  through  them. 
The  riser  runouts,  if  pitched  down  toward  the  riser  as  they  should  be,  are 
sized  the  same  as  the  steam  mains,  and  the  radiator  runouts  are  made  the 
same  as  in  an  up-feed  system. 

In  either  up-feed  or  down-feed  systems  the  returns  are  sized  in  the 
same  manner  and  on  the  same  pressure  drop  basis  as  the  steam  main ;  the 
return  mains  are  taken  from  Columns  0,  R,  U,  X,  or  AA  according  to  the 
drop  used  for  the  steam  main;  and  the  risers  are  sized  by  reading  the 
lower  part  of  Table  8  under  the  column  used  for  the  mains.  The  hori- 
zontal runouts  from  the  riser  to  the  radiator  are  not  usually  increased  on 
the  return  lines  although  there  is  nothing  incorrect  in  this  practice.  The 
same  notes  apply  that  are  given  for  one-pipe  gravity  systems. 

TWO-PIPE  VAPOR  SYSTEMS 

While  many  manufacturers  of  patented  vapor  heating  accessories  have 
their  own  schedules  for  pipe  sizing,  an  inspection  of  these  sizing  tables 
indicates  that  in  general  as  small  a  drop  as  possible  is  recommended.  The 
reasons  for  this  are:  (1)  to  have  the  condensation  return  to  the  boiler  by 
gravity,  (2)  to  obtain  a  more  uniform  distribution  of  steam  throughout 
the  system,  (3)  because  with  large  variation  in  pressure  the  value  of 
graduated  valves  on  radiators  is  destroyed. 

For  small  vapor  systems  where  the  equivalent  length  of  run  does  not 
exceed  200  ft,  it  is  recommended  that  the  main  and  any  runouts  to  risers 
that  may  be  dripped  should  be  sized  from  Column  D,  while  riser  runouts 
not  dripped  and  radiator  runouts  should  employ  Column  /.  The  up-feed 
steam  risers  should  be  taken  from  Column  H.  On  the  returns,  the  risers 
should  be  sized  from  Column  U  (lower  portion)  and  the  mains  from 
Column  U  (upper  portion).  It  should  again  be  noted  that  the  pressure 
drop  in  the  steam  side  of  the  system  is  kept  the  same  as  on  the  return  side 
except  where  the  flow  in  the  riser  is  concerned. 

On  a  down-feed  system  the  main  vertical  riser  should  be  sized  from 
Column  J?,  but  the  down-feed  risers  can  be  taken  from  Column  D  al- 
though it  so  happens  that  the  values  in  Columns  D  and  H  correspond. 
This  will  not  hold  true  in  larger  systems. 

For  vapor  systems  over  200  ft  of  equivalent  length,  the  drop  should  not 
§xceed  ^  Ib  to  J^  Ib,  if  possible.  Thus,  for  a  400  ft  equivalent  run  the 
drop  per  100  ft  shoulcj  be  not  over  J/g  Ib  divided  by  4,  or  %  Ib.  In  this 
case  the  steam  mains  would  be  sized  from  Column  B ;  the  radiator  and 
undripped  riser  runouts  from  Column  I;  the  risers  from  Column  B, 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

because  Column  H  gives  a  drop  in  excess  of  5^  lb.  On  a  down-feed 
system,  Column  B  would  have  to  be  used  for  both  the  main  riser  and  the 
smaller  risers  feeding  the  radiators  in  order  not  to  increase  the  drop  over 
J^2  Ib-  The  return  risers  would  be  sized  from  the  lower  portion  of  Column 
0  and  the  dry  return  main  from  the  upper  portion  of  the  same  column, 
while  any  wet  returns  would  be  sized  from  Column  N.  The  same  pressure 
drop  is  applied  on  both  the  steam  and  the  return  sides  of  the  system. 

Notes  on  Vapor 'Systems 

1.  Radiator  runouts  over  8  ft  long  should  be  increased  one  pipe  size. 

2.  Pitch  of  mains  should  be  not  less  than  y%  in.  in  10  ft. 

3.  Pitch  of  horizontal  runouts  to  risers  and  radiators  should  be  not  less  than  1A  in. 
in  10  ft. 

4.  In  general  it  is  not  desirable  to  have  a  supply  main  smaller  than  2  in.,  and  when  the 
supply  main  is  3  in.  or  over  at  the  boiler  or  pressure  reducing  valve  it  should  be  not  less 
than  2 J^  in.  at  the  far  end. 

5.  When  necessary,  supply  main,  supply  risers,  or  branches  to  supply  risers  should  be 
dripped  separately  into  a  wet  return.    The  drip  for  a  vapor  system  may  be  connected 
into  the  dry  return  through  a  thermostatic  drip  trap. 

VACUUM  SYSTEMS 

Vacuum  systems  are  usually  employed  in  large  installations  and  have 
total  drops  varying  from  J^  to  J^  Ib.  Systems  where  the  maximum 
equivalent  length  does  not  exceed  200  ft  preferably  employ  the  smaller 
pressure  drop  while  systems  over  200  ft  equivalent  length  of  run  more 
frequently  go  to  the  higher  drop,  owing  to  the  relatively  greater  saving  in 
pipe  sizes.  For  example,  a  system  with  1200  ft  longest  equivalent  length 
of  run  would  employ  a  drop  per  100  ft  of  J^  Ib  divided  by  12,  or  J^4  Ib. 
In  this  case  the  steam  main  would  be  sized  from  Column  C,  and  the  risers 
also  from  Column  C  (Column  H  could  be  used  as  far  as  critical  velocity  is 
concerned  but  the  drop  would  exceed  the  limit  of  J^4  Ib).  Riser  runouts, 
if  dripped,  would  use  Column  C  but  if  undripped  would  use  Column  I; 
radiator  runouts,  Column  I;  return  risers,  lower  part  of  Column  S; 
return  runouts  to  radiators,  one  pipe  size  larger  than  the  radiator  trap 
connections. 

Notes  on  Vacuum  Systems 

1.  It  is  not  generally  considered  good  practice  to  exceed  J^-lb  drop  per  100  ft  of 
-equivalent  run  nor  to  exceed  1  Ib  total  pressure  drop  in  any  system. 

2.  Radiator  runouts  over  8  ft  long  should  be  increased  one  pipe  size. 

3.  Pitch  of  mains  should  be  not  less  than  Y%  in,  in  10  ft. 

4.  Pitch  of  horizontal  runouts  to  risers  and  radiators  should  be  not  less  than  }^  in. 
in  10  ft. 

5.  In  general  it  is  not  considered  desirable  to  have  a  supply  main  smaller  than  2  in. 
When  the  supply  main  is  3  in.  or  over,  at  the  boiler  or  pressure  reducing  valve,  it  should 
be  not  less  than  2 J^  in.  at  the  far  end. 

6.  When  necessary,  the  supply  main,  supply  riser,  or  branch  to  a  supply  riser  should 
be  dripped  separately  through  a  thermostatic  trap  into  the  vacuum  return.    A  connec- 
tion should  not  be  made  between  the  steam  and  return  sides  of  a  vacuum  system  withoftt 
interposing  a  thermostatic  trap  to  prevent  the  steam  from  entering  the  return  line. 

7.  Lifts  should  be  avoided  if  possible,  but  when  they  cannot  be  eliminated  they 
should  be  made  in  the  manner  described  in  Chapter  31  under  Up- Feed  Vacuum  Systems. 

538 


CHAPTER  32 — PIPING  FOR  STEAM  HEATING  SYSTEMS 


ATMOSPHERIC  SYSTEMS 

The  sizing  of  the  supply  and  return  piping  on  atmospheric  systems  is 
practically  identical  with  the  sizing  used  for  vacuum  systems  and  the 
same  notes  apply,  except  that  no  lift  can  be  made  in  the  return  line. 

SUB-ATMOSPHERIC  SYSTEMS 

Any  properly  pitched,  correctly  sized  vacuum  system  without  a  lift 
may  be  used  as  a  sub-atmospheric  system  when  the  proper  equipment  is 
substituted  for  the  ordinary  vacuum  pump,  traps,  and  controls.  On  new 
systems  manufacturers  usually  recommend  a  drop  on  the  steam  line  of 
between  J^  and  Y%  Ib  for  the  total  run,  and  suggest  adding  25  ft  to  the 
total  equivalent  length  of  run  to  insure  that  the  steam  gets  through  to  the 
last  radiator. 

The  same  notes  apply  to  these  systems  as  for  vacuum  systems,  except 
that  no  lifts  can  be  made  in  the  returns. 

ORIFICE  SYSTEMS 

The  orifice  systems  can  be  operated  with  any  piping  system  suitable 
for  vacuum  operation,  according  to  experienced  designers.  Because  these 
systems  vary  considerably  in  detail,  it  is  advisable  to  consult  the  manu- 
facturer of  the  particular  system  contemplated  for  recommendations. 

The  same  notes  apply  to  these  systems  as  to  vacuum  systems,  except 
that  lifts  cannot  be  made  in  the  returns  of  orifice  systems  if  a  vacuum 
pump  is  used. 

HIGH  PRESSURE  STEAM 

When  steam  heating  systems  are  supplied  with  steam  from  a  high 
pressure  plant,  one  or  more  pressure-reducing  valves  are  used  to  bring  the 
pressure  down  to  that  required  by  the  heating  system.  It  has  been  con- 
sidered good  practice  to  make  the  pressure  reductions  in  steps  not  to 
exceed  50  Ib  in  each  case.  For  example,  in  reducing  from  100-lb  gage  to 
2-lb  gage,  two  pressure  reducing  valves  would  be  used,  the  first  reducing 
the  pressure  from  100-lb  gage  to  50  Ib  and  the  second  reducing  the  pressure 
from  50-lb  gage  to  2-lb  gage.  Valves  are  available  that  will  reduce  100  Ib 
in  one  step,  and  it  is  questionable  whether  two  valves  are  now  required 
for  initial  pressures  of  150  Ib  or  less. 

The  pressure-reducing  valve,  or  pressure-regulator  as  it  is  sometimes 
termed,  has  ratings  which  vary  200  to  400  per  cent.  Some  of  these 
ratings  are  based  on  arbitrary  steam  velocities  through  the  valve  of 
5,000  to  10,000  fpm  and  it  is  assumed  that  the  valve  when  wide  open  has 
the  same  area  as  the  pipe  on  the  inlet  opening  of  the.  valve.  It  is  well 
known  that  steam  flowing  through  an  orifice  increases  its  velocity  until 
the  pressure  on  the  outlet  side  is  reduced  to  58  per  cent  of  the  absolute 
pressure  on  the  inlet  side,  and  that  with  further  reduction  of  pressure  on 
the  outlet  side  little  change  in  velocity  will  be  obtained.  As  practically 
all  pressure-reducing  valves  used  for  steam  heating  work  lower  the  steam 
pressure  to  less  than  58  per  cent  of  the  inlet  pressures,  only  the  maximum 
velocity  through  such  valves  need  be  considered.  If  it  is  assumed  that 
the  valve,  when  fully  open,  has  an  area  equal  to  that  of  the  inlet  pipe  size, 

539 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


TABLE  10.    CAPACITIES  OF  PRESSURE-REDUCING  VALVES 
(100-LB  GAGE  DOWN  TO  ANY  PRESSURE — 52  LB  OR  LESS) 


INLET  NOMINAL 
PIPE  DIAMETER 
(INCHES) 

POTTNDS  STEAM 

PER  HOXTR 

AT  100-Ln  GAGE 

EQUIVALENT  DIRECT 
RADIATION  SQ  FT 

AT  MLB 

EQUIVALENT  DIRECT 
RADIATION  SQ  FT 
AT  1/3  LB 

1A 

866 

3,464 

2,598 

% 

1,576 

6,304 

4,728 

2,459 

9,836 

7,377 

Ui 

4,263 

17,052 

12,689 

1H 

5,808 

23,232 

17,424 

2 

9,564 

38,256 

28,692 

2^ 

13,623 

54,492 

40,869 

3 

21,041 

84,104 

63,123 

&A 

28,213 

112,852 

84,039 

4 

36,285 

145,140 

108,855 

5 

56,971 

227,884 

170,913 

6 

82,336 

329,344 

247,008 

Formula: 


where 


-        =  pounds  per  hour  passed  by  orifice. 


.4  =  area  of  inlet  pipe,  square  inches. 

V  —  velocity  of  steam  through  orifice  (approximately  870  fps). 

50  ~  70  per  cent  efficiency  of  orifice  less  20  per  cent  for  factor  of  safety, 
144  =  square  inches  in  1  sq  ft. 
3600  =  seconds  in  one  hour. 
3.8  »  cubic  feet  per  pound  at  100-lb  gage. 

that  the  steam  is  flowing  into  a  pressure  less  than  58  per  cent  of  the  initial 
pressure,  that  the  orifice  efficiency  is  approximately  70  per  cent,  and  that 
20  per  cent  more  is  allowed  for  a  factor  of  safety,  then  the  pressure 
reducing  valves  will  have  the  working  capacities  shown  in  Table  10.  If 
the  valve,  when  fully  open,  does  not  give  an  orifice  area  equal  to  that  of 
the  pipe  on  the  inlet  side,  then  the  capacities  will  be  proportional  to  the 
percentage  of  opening  secured,  taking  the  pipe  area  as  100  per  cent. 

Most  exact  regulation  of  pressure  on  steam  heating  systems  is  secured 
from  diaphragm-operated  valves  controlled  by  a  pilot  line  from  the  low 
pressure  pipe,  taken  off  the  low  pressure  main:  at  least  15  ft  from  the 
reducing  valve.  The  reducing  valves  operating  on  the  proportional- 
reduction  principle  will  give  a  variation  of  steam  pressure  on  the  low 
pressure  side  if  the  initial  pressure  varies  between  considerable  limits. 
The  so-called  dead-end  valve  is  used  for  reduced  pressures  where  the  line 
has  not  sufficient  condensing  capacity  at  all  times  to  condense  the  leakage 
that  might  occur  with  the  ordinary  valve.  Single-disc  valves  do  not  give 
as  close  regulation  as  double-disc  valves,  but  the  single  disc  is  preferable 
where  dead-end  valves  are  necessary,  such  as  on  short  runs  to  thermo- 
statically controlled  hot  water  heaters,  central  fan  heating  units  and 
unit  heaters. 

The  correct  installation  (Fig.  2)  of  a  pressure-reducing  valve  includes 
a  pressure-reducing  valve  with  a  gate  valve  on  each  side,  a  by-pass  con- 
trolled by  a  globe  valve,  a  pressure  gage  on  the  low  pressure  side,  and  a 
safety  valve  on  the  low  pressure  main  at  some  point,  usually  within  a 
reasonable  distance  of  the  pressure-reducing  valve.  Pressure-reducing 
valves  should  have  expanded  outlets  for  sizes  greater  than  2  in.  Where 
the  steam  main  is  of  still  larger  diameter  than  the  expanded  outlet,  and  in 

540 


CHAPTER  32 — PIPING  FOR  STEAM  HEATING  SYSTEMS 


cases  where  straight  valves  are  used,  an  increaser  is  placed  close  against 
the  outlet  of  the  valve  to  reduce  the  velocity  immediately  after  passing 
through  the  valve.  Strainers  are  recommended  on  the  inlets  of  all 
pressure-reducing  valves.  A  pressure  gage  may  be  located  on  the  high- 
pressure  line  near  the  valve  if  desired. 

Owing  to  the  large  variation  in  steam  demand  on  the  average  heating 
system,  it  is  generally  advisable  to  use  two  pressure-reducing  valves  con- 
nected in  parallel.  One  valve  should  be  large  enough  for  the  maximum 
load  and  the  other  should  have  a  diameter  approximately  half  that  of  the 
first.  The  smaller  valve  can  be  used  most  of  the  time,  for  it  will  give 
much  better  regulation  than  the  larger  one  on  light  or  normal  loads. 

Less  trouble  from  expansion  leaks  will  occur  when  the  bypass 
valve  is  on  the  same  center  line  as  the  pressure  reducing  valve 

Bypass  (same  size  as  high  fc-^Globe  valve 

pressure  supply  line) N 

Pressure  gage^      jj  Gate  valve  U  /Pressure  gage 

if  desired 


..      f       „  ^^,^,, ...      Low  pressure  steam 

High  pressure  Steam 


Drjp'      strainer  /  i         Gate  valve 


Pressure  reducing  valve      x  Pilot  line 

FIG.  2.    TYPICAL  PRESSURE-REDUCING  VALVE  INSTALLATION 

Control  Valves 

Gate  valves  are  recommended  in  all  cases  where  service  demands  that 
the  valve  be  either  entirely  open  or  entirely  closed,  but  they  should  never 
be  used  for  throttling.  Angle  globe  valves  and  straight  globe  valves 
should  be  used  for  throttling,  as  done  on  by-passes  around  pressure 
reducing  valves  or  on  by-passes  around  traps. 

EXPANSION  IN  STEAM  AND  RETURN  LINES 

Because  all  steam  and  return  lines  expand  and  contract  with  changes 
in  temperature,  provision  should  be  made  for  such  movement.  The 
expansion  in  steam  supply  pipes  is  normally  taken  at  1J^  to  1^  in.  per 
100  ft  and  in  return  lines  at  one-half  or  two-thirds  of  this  amount.  It 
may  be  calculated  accurately  if  the  temperature  rise  and  fall  can  be 
determined  with  reasonable  certainty  (Page  586,  Chapter  34).  The  tem- 
perature at  the  time  of  erection  often  has  a  greater  expansion  effect  on 
piping  than  the  temperature  in  the  building  after  it  has  been  put  into 
service. 

Expansion  may  be  taken  care  of  by  any,  or  all,  of  three  different 
methods,  namely,  (1)  the  spring  in  the  pipe  including  offsets  and  expan- 
sion bends,  (2)  the  turning  of  the  pipe  on  its  threads  and  swing  joints,  and 
(3)  the  use  of  expansion  joints. 

By  the  first  scheme,  which  is  the  most  popular  method  where  space 
permits,  the  pipe  is  offset,  or  broken,  around  rooms  or  corners,  and  is  hung 
so  that  the  spring  in  the  pipe  at  right  angles  to  the  expansion  movement 
is  sufficient  to  absorb  the  expansion*  If  conditions  do  not  lend  themselves 
to  this  treatment,  regular  expansion  bends  of  the  U  or  offset  type  may  be 
used.  In  tight  places  such  as  pipe  tunnels  the  expansion  joint  is  pre- 
ferable. 

541 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


On  riser  runouts  and  radiator  runouts  the  swing  joint  is  used  almost 
without  exception.  On  high  vertical  risers  the  pipes  may  be  reversed 
every  five  to  ten  stories;  that  is,  the  supply  is  carried  over  to  the  adjacent 
return  riser  location  and  the  return  riser  is  run  over  to  the  former  supply 
riser  location,  thus  making  horizontal  offsets  in  each  line.  Corrugated 
copper  expansion  joints  also  are  used  on  risers  but  must  be  made  acces- 
sible in  case  future  replacement  becomes  necessary. 

EXPANSION  BENDS 

The  calculation  of  the  distance  required  for  offsets  and  the  size  of 
expansion  bends  necessary  to  absorb  a  given  amount  of  expansion  leads 
into  complicated  formulae  and  is  a  subject  of  controversy.  It  seems  to 
have  been  demonstrated,  however,  that  the  shape  of  the  bend,  the  radius 
used,  the  relative  amounts  of  straight  and  curved  pipe  in  a  bend,  and  the 


Fitting  offset 


A+B+OL 

U  bend  with 

4  fittings 


Offset  bend 


Radial  U  bend 
rr; -^  Circle  bend 

.Offset  U  bend 

FIG.  3.    MEASUREMENT  OF  L  ON  VARIOUS  PIPE  BENDS  AND  OFFSETS 
FOR  ABSORBING. EXPANSION 

type  of  bend  have  little  bearing  on  the  amount  of  expansion  for  which 
they  will  safely  provide.  The  size,  weight  and  material  of  the  pipe  and  the 
length  of  all  of  the  pipe  in  the  bend,  or  even  in  the  offset,  have  a  bearing 
on  its  capacity  to  absorb  expansion  without  straining  the  pipe  material 
beyond  the  safe  working  stress.  In  Fig.  3  typical  pipe  bends  and  offsets 
for  absorbing  expansion  are  shown.  The  lengths  L  are  those  which  are 
used  in  determining  the  stress  in  the  pipe. 

Fig.  4  shows  a  set  of  curves  for  standard  weight  steel  pipe  bends  from 
which  the  approximate  amount  of  pipe  L  (Fig.  3)  for  each  pipe  size  may 
be  determined  from  the  amount  of  expansion  movement  that  must  be 
absorbed.  These  curves  are  such  that  the  maximum  fiber  stress  in  any 
part  of  the  bend  will  not  be  over  16,000  Ib  per  square  inch.  Since  12,000 
Ib  per  square  inch  is  considered  to  be  a  maximum  working  fiber  stress  in 
wrought  iron  pipe,  an  additional  33  J4  per  cent  must  be  added  to  the 
length  of  this  type  of  pipe. 

The  amount  of  expansion  can  be  doubled  for  a  bend  if  the  bend  is  cold 
sprung  for  one-half  of  the  expansion  movement.  In  other  words,  if  the 
bend  is  erected  with  the  main  pipe  cut  short  one-half  of  the  expected 
expansion  and  the  bend  is  then  sprung  open  to  meet  the  shortened  pipe, 

542 


CHAPTER  32 — PIPING  FOR  STEAM  HEATING  SYSTEMS 


23456 
EXPANSION  ABSORBED  IN  INCHES 


FIG.  4.    CURVES  GIVING  LENGTH  L  OF  BEND  OF  OFFSET  NECESSARY  TO 
ABSORB  EXPANSION  (WITHOUT  COLD  SPRING) 

the  expansion  in  the  main  will  first  allow  the  bend  to  go  back  to  its  neutral 
point  and  then  will  compress  the  bend  an  equal  distance  beyond  the 
neutral  point,  thus  securing  a  doubled  capacity.  Generally  only  a  portion 
of  the  cold  spring  is  considered  as  being  effective,  owing  to  the  difficulties 
of  erecting  the  bends  with  sufficient  exactitude  in  the  length  of  the  main 
line  and  the  difficulty  of  cold  springing.  See  additional  material  on  pipe 
expansion  in  Chapter  34. 

PIPING  CONNECTIONS  AND  DETAILS 

Piping  connections  may  be  classified  into  two  groups:  first,  those 
suitable  for  any  system  of  steam  heating;  second,  those  devised  for  certain 
systems  which  cannot  be  satisfactorily  applied  to  any  other  type.  There 
are  also  various  details  that  apply  to  piping  on  the  steam  side  which 
cannot  be  used  on  the  returns.  An  installation  that  is  designed  and  sized 
correctly  and  installed  with  care  may  be  rendered  defective  by  the  use  of 
improper  connections,  such  as  runouts  that  do  not  allow  for  expansion, 
thermostatic  traps  unprotected  from  scale,  pressure-reducing  valves 
without  strainers,  and  lack  of  drips  at  required  points. 

BOILER  CONNECTIONS 
Supply 

Boiler  headers  and  connections  have  the  largest  sizes  of  pipe  used  in  a 
system.  Cast-iron,  horizontal-type,  low  pressure  heating  boilers  usually 
have  several  tapped  outlets  in  the  top,  the  manufacturers  recommending 
their  use  in  order  to  reduce  the  velocity  of  the  steam  in  the  vertical  up- 
takes from  the  boiler  and  to  permit  entrained  water  to  return  to  the 
boiler  instead  of  being  carried  over  into  the  steam  main  where  it  must  be 
cared  for  by  dripping.  Steel  heating  boilers  usually  are  equipped  with 
only  one  steam  outlet  but  many  engineers  believe  that  better  results  are 
obtained  by  specifying  that  such  boilers  have  two.  The  second  outlet, 
usually  located  3  or  4  ft  back  of  the  regular  one,  reduces  the  velocity 
5Q  per  cent  in  the  steam  uptake. 

Fig.  5  shows  a  type  of  boiler  connection  that  was  used  for  many  years 
and  one  with  which  some  boilers  are  now  piped.  The  uptakes  are  carried 
as  high  as  possible,  turned  horizontally  and  run  out  to  the  side  of  the 
boiler  and  then  are  connected  together  into  the  main  boiler  runout  which 

543 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

drops  into  the  top  of  the  boiler  header  through  a  boiler  stop  valve.  No 
drips  are  provided  on  this  type  of  runout  except  a  very  small  one  which 
is  sometimes  installed  on  the  boiler  side  of  the  stop  valve.  Fig.  6  shows  a 
type  of  boiler  connection  which  is  regarded  as  superior  to  that  shown  in 
Fig.  5  and  which  is  the  type  illustrated  in  the  system  diagrams  in  Chapter 
31.  This  type  is  similar  to  that  shown  in  Fig.  5  except  that  the  horizontal 
branches  from  the  uptakes  are  connected  into  the  main  boiler  runout,  and 
the  steam  is  carried  toward  the  rear  of  the  boiler.  The  branch  to  the 
building  or  boiler  header  is  taken  off  behind  the  last  horizontal  boiler  con- 
nection. At  the  rear  end  of  this  main  runout,  a  large  size  drip,  or  balance 
pipe,  is  dropped  down  into  the  boiler  return,  or  into  the  top  of  the  Hart- 
ford Loop,  which  is  described  in  a  following  paragraph.  As  a  result,  any 
water  carried  over  from  the  boiler  follows  the  direction  of  steam  flow 


Reducing  ell 
\ 

Uptake  - 

Main  runout  to 

building  or 

header 

•Drip  and  balance  pipe 
Water  line 
Hartford  return  connection 

Main  wet  return 


FIG.  5.    OLD  STYLE  STANDARD  BOILER 
CONNECTIONS 


FIG.  6.    APPROVED  METHOD  OF  BOILER 
CONNECTIONS 


toward  the  rear  and  is  discharged  into  the  rear  drip,  or  balance  pipe, 
without  being  carried  over  into  the  system. 

Return 

Cast-iron  boilers  are  generally  provided  with  return  tappings  on  both 
sides,  but  steel  boilers  often  are  equipped  with  only  one  return  tapping. 
A  boiler  with  side  return  tappings  will  usually  have  a  more  effective  cir- 
culation if  both  tappings  are  used.  Check  valves  generally  should  not  be 
used  on  the  return  connection  to  steam  heating  boilers  because  they  are 
not  always  dependable  inasmuch  as  a  small  piece  of  scale  or  dirt  lodged 
on  the  seat  will  hold  the  tongue  open  and  make  the  check  useless.  These 
valves  also  offer  a  certain  amount  of  resistance  to  the  returns  coming  back 
to  the  boiler,  and  in  gravity  systems  will  raise  the  water  line  in  the  far 
end  of  the  wet  return  several  inches3.  However,  if  check  valves  are 
omitted  and  the  steam  pressure  is  raised  with  the  boiler  steam  valve 
closed,  the  water  in  the  boiler  will  be  blown  out  into  the  return  system 
with  the  accompanying  danger  of  boiler  damage.  These  objections  are 
largely  overcome  with  the  Hartford  return  connection. 


•Sec  method  of  calculating  height  above  water  line  for  gravity  one-pipe  systems  in  Chapter  31. 

544 


CHAPTER  32 — PIPING  FOR  STEAM  HEATING  SYSTEMS 


Hartford  Return  Connection 

In  order  to  prevent  the  boiler  from  losing  its  water  under  any  circum- 
stances, the  use  of  the  Hartford  Connection,  or  the  Underwriters  Loop, 
is  recommended.  Fig.  7  shows  this  connection  for  both  single  boiler  and 
two-boiler  installations.  By  balancing  the  column  of  water  in  the  loop 
against  the  steam  pressure,  the  water  cannot  be  blown  out  of  the  loop 
whatever  the  relative  pressure  conditions  in  the  boiler,  steam  lines,  or 
return  lines.  This  balancing  is  done  by  raising  the  return  to  approxi- 
mately the  normal  water  line  of  the  boiler,  looping  it  back  to  the  boiler 
inlet  and  connecting  the  top  of  this  loop  by  means  of  a  balance  pipe  with 
the  steam  runout  from  the  boiler.  It  is  important  that  this  balance  pipe 
be  connected  into  the  boiler  steam  line  on  the  boiler  side  of  all  valves. 


rTHESE  CONNECTIONS  SHOULD  &E  AS  SHO&T  AS  PRACTICAL  USING  AS  FEW-, 
\TUQNS  AS  POSSIBLE.    THE.  SIZES  SHOULD  PBEFECABLY  BE  NOT 
SMALLER  THAN  GlVEM  IN  TABLE  BELOW. 

GfcATE  AREA                PIPE  SIZE 
4  SQ.  FEET  OS  LESS 11// 


THESE  CONNECTIONS  TO  &E 
'  TAKEN  OFF  BETWEEN  STOP 
'  VALVES  (IF  ANY)  AMD  BOILEfc 


THESE  PIPES  MAY  BE  ANY  SIZE  COUSlDE&ED  PHOPEa  FOB  I 
BOILERS  AWD  LESS  THAN  STEAM  CONWECTIQUS  IN  TABLE  ABOVE." 

FIG.  7.    THE  HARTFORD  RETURN  CONNECTION 

Theoretically,  the  top  of  the  loop  should  be  at  the  normal  boiler  water 
line  but  since  this  installation  often  causes  trouble  from  water  hammer  in 
the  top  of  thejloop,  this  top  is  usually  made  2  in.  below  the  normal  boiler 
water  line  to  keep  the  horizontal  pipe  at  the  top  submerged  under  all 
normal  conditions.  It  is  important  that  this  top  of  the  loop  be  made  with 
the  shortest  possible  horizontal  pipe,  a  close  nipple  being  employed. 

Sizing  Boiler  Connections 

Little  authentic  information  is  available  on  the  sizing  of  boiler  runouts 
and  steam  headers.  Although  many  engineers  prefer  an  enlarged  steam 
header  to  serve  as  additional  steam  storage  space,  there  ordinarily  is  no 
sudden  demand  for  steam  in  a  steam  heating  system  except  during  the 
heating-up  period,  at  which  time  a  large  steam  header  is  a  disadvantage 
rather  than  an  advantage.  The  boiler  header  may  be  sized  by  first  com- 
puting the  maximum  load  that  must  be  carried  by  any  portion  of  the 
header  under  any  conceivable  method  of  operation,  and  then  applying 
the  same  schedule  of  pipe  sizing  to  the  header  as  is  used  on  the  steam 
mains  for  the  building.  The  horizontal  runouts  from  the  boiler,  or  boilers, 
may  be  sized  by  calculating  the  heaviest  load  that  will  be  placed  on  the 
boiler  at  any  time,  and  sizing  the  runout  on  the  same  basis  as  the  building 
mains.  The  difference  in  size  between  the  vertical  uptakes  from  the 

545 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


6,000  sq  ft 


2,000  sq  ft 


LI 


8,000  sq  ft 


8,000  sq  ft 


8,000  sq  ft 


FIG.  8.    BOILER  STEAM  HEADER  AND  CONNECTIONS 


boiler  and  the  horizontal  main  or  runout  is  compensated  for  by  the  use  of 
reducing  ells  (Figs.  5  and  6). 

The  following  example  illustrates  the  sizing  of  the  boiler  connections 
shown  in  Fig.  8. 

Example  4-  Determine  the  size  of  boiler  steam  header  and  connections  (Fig.  8)  if 
there  are  three  boilers,  two  to  carry  50  per  cent  of  the  load  each,  and  the  third  to  be  used 
as  a  spare.  The  steam  mains  are  based  on  J^-lb  drop  per  100  sq  ft  of  equivalent  direct 
radiation  (EDR). 


Solution: 


Size  of  Boiler  Header 


WHEN 
OPERATING 
ON  BOILIRS 

LOAD  ON  VAEIOUS  PORTIONS  OF  HEADER 

MAXIMUM 
LOAD 

A 

B 

C 

D 

E 

F 

Nos.  1  and  2 
Nos.  2  and  3 
Nos.  3  and  1 

6000 
6000 
6000 

0 
6000 
0 

2000 
8000 
2000 

4000 
2000 
2000 

3000 
3000 
3000 

3000 
3000 
3000 

6000 
8000 
6000 

Max.  Load 

6000 

6000 

8000 

4000 

3000 

3000 

8000 

8000  sq  ft  @  H  lb  per  100  ft  *  6  in.  main.    (See  Table  7.) 

Size  of  Boiler  Runouts 


The  three  runouts 
GI.  G^t  Gs 


«  2667  sq  ft  each 

«i> 


4  in,  pipe. 


i,  JSTa,  H8  =  2667  sq  ft  each 
i,  /a,  J*  =  5333  sq  ft  each 
i,  Ki,  Kt  =  8000  sq  ft  each 


£  lb  per  100  ft 

Jg  lb  per  100  f  t  =  4  in.  pipe4  (See  Table  7). 
^  lb  per  100  f  t  «  5  in.  pipe4  (See  Table  7), 
J$  lb  per  100  ft  «  6  in.  pipe4  (See  Table  7), 


The  uptakes  from  the  boiler  probably  would  be  6  in.  pipe  with  a  6  in.  X  4  in.  reducing 
ell  at  top. 

Return  connections  to  boilers  in  gravity  systems  are  made  the  same 
size  as  the  return  main  itself.    Where  the  return  is  split  and  connected  to 

Wote.  —  As  JCi,  Ka,  JCi  all  carry  8000  sq  ft  and  are  6  in.  pipe,  the  whole  runout  including  Ji,  /a  and  /* 
would  be  made  6  in.  pipe,  also. 

546 


CHAPTER  32 — PIPING  FOR  STEAM  HEATING  SYSTEMS 


two  tappings  on  the  same  boiler,  both  connections  are  made  the  full  size 
of  the  return  line.  Where  two  or  more  boilers  are  in  use,  the  return  to 
each  may  be  sized  to  carry  the  full  amount  of  return  for  the  maximum  load 
which  that  boiler  will  be  required  to  carry.  Where  two  boilers  are  used, 
one  of  them  being  a  spare,  the  full  size  of  the  return  main  would  be 
carried  to  each  boiler,  but  if  three  boilers  are  installed,  with  one  spare,  the 
return  line  to  each  boiler  would  require  only  half  of  the  capacity  of  the 
entire  system,  or,  if  the  boiler  capacity  were  more  than  one-half  the  entire 
system  load,  the  return  would  be  sized  on  the  basis  of  the  maximum 
boiler  capacity.  As  the  return  piping  around  the  boiler  is  usually  small 
and  short,  it  should  not  be  sized  to  the  minimum. 

With  returns  pumped  from  a  vacuum  or  receiver  return  pump,  the  size 
of  the  line  may  be  calculated  from  the  water  rate  on  the  pump  discharge 
when  it  is  operating,  and  the  line  sized  for  a  very  small  pressure  drop,  the 
size  being  obtained  from  the  Chart  for  Friction  Losses  for  Various  Rates 
of  Flow  of  Water,  Fig.  3,  Chapter  35.  The  relative  boiler  loads  should  be 
considered,  as  in  the  case  of  gravity  return  connections. 

Radiator  Connections 

Radiator  connections  are  important  on  account  of  the  number  of 
repetitions  which  occur  in  every  heating  installation.  They  must  be 
properly  pitched  and  they  must  be  arranged  to  allow  not  only  for  move- 
ment in  the  riser  but,  in  frame  buildings,  for  the  shrinkage  of  the  building. 
In  a  three  story  building  this  sometimes  amounts  to  1  in.  or  more.  The 
simplest  connection  is  that  for  the  one-pipe  system  where  only  one  radia- 
tor connection  is  necessary.  Where  the  radiator  runouts  are  located  on 
the  ceiling  or  under  the  floor,  sufficient  space  usually  is  available  to  make 
a  good  swing  joint  with  plenty  of  pitch,  but  where  the  runouts  must  come 
above  the  floor  the  vertical  space  is  small  and  the  runouts  can  project  out 
into  the  room  only  a  short  distance.  Fig.  9  illustrates  two  satisfactory 
methods  of  making  runouts  on  a  one-pipe  gravity  air  vent  system  of 
either  the  up-feed  or  down-feed  type,  the  runout  below  the  floor  being 
indicated  in  full  lines  and  the  runout  above  the  floor  in  dotted  lines. 
Sometimes  it  is  necessary  to  set  a  radiator  on  pedestals,  or  to  use  high 
legs,  in  order  to  obtain  sufficient  vertical  distance  to  accommodate  above- 
the-floor  runouts.  Particular  attention  must  be  given  to  the  riser  expan- 
sion as  it  will  raise  the  runout  and  thereby  reduce  the  pitch. 

Similar  connections  for  a  two-pipe  system  of  the  gravity  air  vent  type 
are  illustrated  in  Fig.  10  for  the  old  steam  type  radiator.  If  the  water 
type  is  used,  the  supply  tapping  is  at  the  top  instead  of  at  the  bottom,  the 
runouts  otherwise  remaining  as  shown  in  Fig.  10.  A  satisfactory  type 
of  radiator  connection  for  atmospheric,  vapor,  vacuum,  sub-atmos- 
pheric, and  orifice  systems  of  both  the  up-feed  and  down-feed  types  is 
shown  in  Fig.  11. 

While  short  radiators,  not  exceeding  8  to  10  sections,  may  be  supplied 
and  returned  from  the  same  end  as  indicated  in  Fig.  12,  the  top-an- 
bottom-opposite-end  method  is  to  be  preferred  in  all  cases  where  it  can  be 
used.  On  down -feed  systems  of  the  atmospheric,  vapor,  vacuum,  sub- 
atmospheric,  and  orifice  types,  the  bottom  of  the  supply  riser  must  be 
dripped  into  the  return  somewhat  as  illustrated  in  Fig.  13.  On  up-feed 
systems  of  the  vapor  and  atmospheric  types,  where  radiators  in  the 

547 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


;  VALVE 


-•Runout  below  floor 
PLAN 


L 


Globe 

Runout  valve 
above 
(floor       i| 


Hot 


lyFloor 


Runout  below  floor 
ELEVATION 


FIG.   9.     TYPICAL  ONE-PIPE  RADIATOR 
CONNECTIONS  (UP-FEED  OR  DOWN-FEED) 


KCENTQlC 

BUSHING 


FIG.  12.    TOP  AND  BOTTOM  RADIATOR 

CONNECTIONS  FROM  UP-  OR  DOWN-FEED 

RISERS.      (NOT    TO  EXCEED  8  TO  10 

SECTIONS.) 

Note.—  Suitable  for  up-feed  or  down-feed  at- 
mospheric, vapor,  vacuum,  sub-atmospheric,  and 
orifice  systems.  Opposite  end  connections  always 

preferable. 


WATEQ-TYPE 
/  CA01ATOQ 


ECCENTQlC 

^  BUSHIMG 
•TUAP 


-TBAP 


FIG.  10.    CONNECTIONS  TO  STEAM-TYPE 

RADIATOR   FOR   TWO-PIPE    GRAVITY 

SYSTEM,  UP-FEED  OR  DOWN-FEED 

Note. — Steam-type  radiators  should  not  be  used 
on  any  except  gravity  one-pipe  and  gravity  two- 
pipe  systems. 


FIG.  13.    TOP  AND  BOTTOM  OPPOSITE  END 

RADIATOR  CONNECTIONS  WITH  HEEL  OF 

DOWN-FEED  RISER  DRIPPED  INTO 

DRY  RETURN 

Note. — Suitable  for  down-feed  only.  For  at- 
mospheric, vapor,  vacuum,  sub-atmospheric,  and 
orifice  systems. 


FIG.  11.    TOP  AND  BOTTOM  OPPOSITE  END 

RADIATOR  CONNECTIONS  FROM  UP 

OR  DOWN-FEED  RISERS 

Not«.t — Suitable  for  up-feed  or  down-feed  at- 
mospheric,  vapor,  vacuum,  sub-atmospheric  and 

orifice  systems. 


FIG,  14.    CONNECTIONS  TO  RADIATOR  HUNG 

ON  WALL 

Note.. — For  up-feed  with  radiators  below  level 
of  steam  main.  For  atmospheric  and  vapor  systems. 
JNot  suitable  for  vacuum,  sub-atmospheric,  or 
orifice  systems. 


548 


CHAPTER  32 — PIPING  FOR  STEAM  HEATING  SYSTEMS 


basement  are  located  below  the  level  of  the  steam  main,  the  drop  to  the 
radiator  is  dripped  into  the  wet  return  and  an  air  line  is  used  to  vent  the 
return  radiator  connection  into  an  overhead  return  line,  as  illustrated  in 
Fig.  14.  When  the  radiator  stands  on  the  floor  below  the  main,  the  drip 
on  the  steam  branch  down  to  the  radiator  may  be  omitted  if  an  overhead 
valve,  as  shown  in  Fig.  15,  is  used.  This  method  is  also  suitable  for 
vacuum,  sub-atmospheric,  and  orifice  systems. 

Convector  Connections 

Convectors  often  are  installed  without  control  valves,  a  damper  being 
used  to  shut  off  the  flow  of  air  to  retard  the  heat  transfer  from  the  con- 
vector  even  though  it  is  still  supplied  with  steam.  The  piping  connec- 
tions for  a  convector  with  the  inlet  and  outlet  at  the  same  end  are  shown 
in  Fig.  16.  There  is  no  valve  on  the  steam  side  but  there  is  a  thermostatic 
trap  on  the  return.  The  damper  for  control  is  shown  immediately  above 
the  convector.  This  piping  is  suitable  for  atmospheric,  vapor,  vacuum, 
sub-atmospheric,  and  orifice  systems  of  the  up-feed  type.  A  similar  unit 
with  connections  on  opposite  ends  and  suitable  for  the  same  systems  is 
shown  in  Fig.  17.  This  unit  has  no  damper  but  requires  a  valve  on  the 
steam  connection  for  control.  When  valves  must  be  located  so  as  to  be 
accessible  from  the  supply  air  grille,  the  arrangement  usually  takes  the 
form  indicated  in  Fig.  18.  Convectors  with  damper  control,  installed  in 
cabinets  or  under  window  sills,  usually  are  connected  as  shown  in  Fig.  19. 
A  convector  located  in  the  basement  and  supplying  air  to  a  room  on  the 
floor  above  may  be  piped  as  pictured  in  Fig.  20  for  all, systems  except 
gravity  one-pipe  or  two-pipe  systems. 

Vapor  systems  with  heating  units  in  the  basement  where  the  returns 
are  wet  would  be  treated  as  in  Fig.  21.  Similar  heating  units  where  a  dry 
return  is  available  would  be  connected  as  shown  in  Fig.  22.  If  the  dry 
return  were  on  a  vacuum,  atmospheric,  sub-atmospheric  or  orifice  system, 
the  treatment  would  be  identical. 

Pipe  Coil  Connections 

Pipe  coils,  unless  coupled  in  a  correct  manner,  often  give  trouble  from 
short  circuiting  and  poor  circulation.  The  method  of  connecting  shown 
in  Fig.  23  is  suitable  for  atmospheric,  vapor,  vacuum,  sub-atmospheric, 
and  orifice  systems. 

Indirect  Air  Heater  Connections 

Heating  units  for  central  fan  systems  have  simple  connections  on  the 
steam  side.  The  steam  main  is  carried  into  the  fan  room  and  has  a 
single  branch  tapped  off  for  each  row  of  heating  units.  Each  of  these 
main  branches  is  split  into  as  many  connections  as  need  be  made  to  each 
row,  governed  by  the  number  of  stacks  and  the  width  of  the  stacks.  Each 
stack  must  have  at  least  one  steam  connection,  and  wide  stacks  are  more 
evenly  heated  with  two  steam  connections,  one  at  each  end. 

The  piping  shown  in  Fig.  24  is  for  small  stacks  and  has  the  steam  con- 
nected at  only  one  end.  On  the  return  side  all  of  the  returns  are  collected 
together  through  check  valves  and  are  passed  through  blast  traps  which 
are  connected  to  the  vacuum  return  or  to  an  atmospheric  return.  The  air 

549 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


550 


CHAPTER  32 — PIPING  FOR  STEAM  HEATING  SYSTEMS 


INDIRECT  QAOATOG. 


-Thermostatic  trap 


DBYBETUBW 
'SUPPLY  MAIN 


FIG.  22.    TYPICAL  PIPING  CONNECTIONS 

TO   INDIRECT   RADIATORS  WITH 

DRY  RETURN 


Full  size  of  tapping  - 

Reducing  ell 


"o  return  line 
beyond  blast  traps 

To  blast  trap 


Check  valve 


Note. — Suitable  for  atmospheric,  vapor,  vacuum,          FlG.  25.      HEATING  UNIT  RETURN  CON- 
sub-atmospheric,  and  orifice  systems.  NECTION  WITH  SEPARATE  AlR  LlNE 


Ill 

1 

Jt 

TYPICAL  CONHKTIONS  TO  MANIFOLD 
COILS  OF  MOT  OVEE  S  PPES 

,                                                                                                                  e 

TYPICAL  CONNECTIONS  TO  MA.N1FOID 
CO\LS  HW1M6  MORE  THAN  6  PPES. 


FIG.  23.    TYPICAL  PIPE  COIL  CONNECTIONS 

Note. — Suitable  for  up-feed  or  down-feed.    For  atmospheric,  vapor,  vacuum,  sub-atmospheric,  and 
orifice  systems. 


BLAST  HEATERS  , 


SUPPLY  AND  RETURN  CONNECTIONS  TO  BLAST  COILS 
FOR  VACUUM  SYSTEM  US)H<5  BLAST  TRAP  OM  EACH  TIER, 

FIG.  24,    CONNECTIONS  FOR  HEATING 
UNITS  OF  CENTRAL  FAN  SYSTEMS 


•STUNNER 
TBAP 


FIG,  26.    TYPICAL  CONNECTIONS  TO 

CENTRAL    FAN    SYSTEM    HEATING 

UNITS  EXCEEDING  12  SECTIONS 


Note. — Suitable  for  atmospheric  and  vacuum  Note. — Suitable  for  vacuum  and  atmospheric 

systems,  systems. 

551 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


from  the  stacks,  in  the  case  illustrated,  passes  up  into  a  small  air  line  and 
through  a  thermostatic  trap  into  a  line  connecting  into  the  return  beyond 
the  blast  trap.  It  is  important  to  use  a  nipple  the  full  size  of  the  outlet 
tapping  on  the  stack  and  to  reduce  the  pipe  size  to  the  normal  return  size 
required,  by  the  use  of  a  reducing  ell,  as  indicated  in  Fig.  25. 

Where  the  stacks  contain  some  thirteen  or  more  sections,  an  auxiliary 
air  tapping  is  made  to  the  lower  portion  of  one  of  the  middle  sections,  in 
the  manner  illustrated  in  Fig.  26,  to  prevent  air  collecting  at  this  point. 
Thermostatic  control  as  applied  to  such  heating  units  in  modern  practice 


"~wTr 'W\  y^: gr~"*y 


Blast  trap  with 

thermostatic 

bypass 


urn  from  lower  tier 

-   —  —   —  —    —    _    «.  _    _....._ 

^Return  from  upper  tier 


To  vacuum  pump 


FIG.  27.    TYPICAL  PIPING  FOR  ATMOSPHERIC  AND  VACUUM  SYSTEMS  WITH 
THERMOSTATIC  CONTROL  (CENTRAL  FAN  SYSTEM) 

consists  of  a  thermostatic  valve  located  in  each  main  branch  from  the 
steam  line  so  that  each  valve  will  open  or  close  a  complete  row  of  stacks 
across  the  entire  face  of  the  heating  unit.  The  stack  closest  to  the  fresh 
air  intake  is  not  usually  equipped  with  a  control  valve.  Steam  is  fur- 
nished continually  to  this  coil  to  prevent  freezing,  and  only  the  supply 
pipe  is  equipped  with  a  gate  valve.  In  this  case  no  particular  attention 
need  be  paid  to  the  method  of  connecting  the  returns,  that  is,  they  do  not 
need  to  be  connected  in  parallel  with  the  steam  connections  but  may  be 
hooked  together  in  any  convenient  manner.  The  arrangement  shown  in 

552 


CHAPTER  32 — PIPING  FOR  STEAM  HEATING  SYSTEMS 


Fig.  27  is  satisfactory.  A  detail  of  the  arrangement  where  a  connection  is 
made  with  a  stack  is  shown  in  Fig.  28.  It  is  essential  to  have  a  check 
valve  on  each  individual  stack  to  prevent  reverse  flow  when  the  ther- 
mostatic valve  in  the  steam  line  closes  off  and  a  partial  vacuum  is  pro- 
duced in  the  stack.  The  end  of  the  steam  main  also  should  be  dripped  as 
indicated  in  Fig.  27. 

If  the  separate  air  line  is  used  as  shown  in  Fig.  24,  the  blast  traps  may 
be  supplied  without  thermostatic  by-passes  but  if  the  piping  is  arranged 
as  shown  in  Figs.  26  or  27,  the  blast  traps  must  be  supplied  with  the 
thermostatic  by-passes  to  permit  the  passage  of  the  air. 

PIPE  SIZING  FOR  INDIRECT  HEATING  UNITS 

Pipe  connections  and  mains  for  indirect  heating  units  are  sized  in  a 
manner  similar  to  radiators,  but  the  equivalent  direct  radiation  must  be 
ascertained  for  each  row  of  heating  unit  stacks  and  then  must  be  divided 
into  the  number  of  stacks  constituting  that  row  and  into  the  number  of 
connections  to  each  stack. 

x  6P  ><_  to_r  ^      Q  x  fa  - 

55.2  X~ 240  "220.8 

where 
EDR 


j?DR  = 


(3) 


equivalent  direct  radiation,  square  feet. 
Q  =  volume  of  air,  cubic  feet  per  minute. 

/e  =  the  temperature  of  the  air  entering  the  row  of  heating  units  under  con- 
sideration, degrees  Fahrenheit. 

t\  =  the  temperature  of  the  air  leaving  the  row  of  heating  units  under  considera- 
tion, degrees  Fahrenheit. 
60  =  the  number  of  minutes  in  one  hour. 
55.2  =  the  number  of  cubic  feet  of  air  heated  1  F  by  1  Btu. 
240  =  the  number  of  Btu  in  1  sq  ft  of  EDR. 

Example  5.  Assume  that  the  heating  units  shown  in  Fig.  27  are  handling  50,000  cfm 
of  air  and  that  the  rise  in  the  first  row  is  from  0  to  40  F,  in  the  second  row  from  40  to 
65  F,  and  in  the  third  row  from  65  to  80  F.  What  is  the  load  in  EDR  on  each  supply 
and  return  connection? 


Solution.    For  row  1, 

j 
For  row  2, 

For  row  3, 


R 


50,000  X  (40  -  0)    _ 
220.8 

50,000  X  (65  -  4 


so 

sq 


"220.8 


R  =  50,000  X  (80  -  65) 


5661  sq  ft. 


3397  sq  ft. 


220.8 

Each  row  of  heating  units  consists  of  four  stacks  and  each  stack  has  two  connections 
so  that  the  load  on  each  stack  and  each  connection  of  the  stack  is  as  follows: 


Row 

TOTAL  LOAD 
(EDR) 

STACK  LOAD* 
(EDR) 

CONNECTION  LoADb 
(EDR) 

1 

9058 

2265 

2265  or  1132 

2 

5661 

1415 

1415  or    708 

3 

3397 

849 

849  or    425 

quarter  of  total  row  load. 
bQne  half  of  stack  load  if  two  steam  connections  are  made;  otherwise,  same  as  stack  load. 

553 


AMERICAN  SOCIETY  of  HEATING  drid  VENTILATING  ENGINEERS  GUIDE,  1935 


Reducing  el! 


To  blast  trap 


FIG.  28.  HEATING  UNIT  RETURN 
CONNECTION  WITHOUT  SEPARATE 
AIR  LINE  (CENTRAL  FAN  SYSTEM) 


r 


TO  flND  LENGTHC  -  MULTIPLY  A, 
BY  CONSTANT  FOQ  ASlQLE  B>. 


FIG.  32,    CONSTANTS  FOR  DETERMINING 
PROPER  LENGTH  OF  OFFSET  PIPE 


45°  ELI 


ACCEPTABLE  METHOD  PGSFEGED  METHOD 

FIG.  33.    ACCEPTABLE  AND  PREFERRED 

METHODS  OF  TAKING  BRANCH 

FROM  MAIN 


FIG.  29.  METHOD 

OF  DRIPPING  MAIN  WHERE 

IT  RISES  TO  HIGHER  LEVEL 

/f.  —  Suitable    for    vapor    and    atmospheric 


systems. 


FIG.  30.   LOOPING 

MAIN  AROUND 
BEAM 


OQT  POCKET 


FIG.  34.     DIRT 

POCKET 
CONNECTION 


TRAP 


L  SUPPLY  MA\N 


FIG.  31.   LOOPING  DRY 

RETURN  MAIN  AROUND 
OPENING 

ATo^.— -Suitable  for  any  dry  return  line  and  any 

return  line' carrying  air*  . 


REDUCING  CQUPUNQ-x, 

FIG.  35,    DRIPPING  END  OF 
MAIN  INTO  WET  RETURN 

,2Vote.-— Suitable-  for*  vapor  systems. 


554 


CHAPTER  32 — PIPING  FOR  STEAM  HEATING  SYSTEMS 


DBY  PETUBU- 

• 

FIG.  36.     DRIPPING  END  OF  MAIN  INTO  FIG.  37.    DRIPPING  HEEL  OF  RISER  INTO 
DRY  RETURN.    (A  GATE  VALVE  is  DRY  RETURN.  (A  GATE  VALVE  is 

RECOMMENDED  AT  THE  INLET  RECOMMENDED  AT  THE  INLET 

SIDE  OF  THE  TRAP)  SIDE  OF  THE  TRAP) 

The  pipe  sizes  would  then  be  based  on  the  length  of  the  run  and  the  pressure  drop 
desired,  as  in  the  case  of  radiators.  It  generally  is  considered  desirable  to  place  the  in- 
direct heating  units  on  a  separate  system  and  not  on  supply  or  return  lines  connected  to 
the  general  heating  system. 

DRIPPING 

Any  steam  main  in  any  type  of  steam  heating  system  may  be  dropped 
to  a  lower  level  without  dripping  if  the  pitch  is  downward  with  the  steam 
flow.  Any  steam  main  in  any  heating  system  can  be  elevated  if  dripped 
(Fig.  29).  Steam  mains  also  may  be  run  over  obstructions  without  a 
change  in  level  if  a  small  pipe  is  carried  below  the  obstruction  to  care  for 
the  condensation  (Fig,  30).  Return  mains  may  be  carried  past  doorways 
or  other  obstructions  by  using  the  scheme  illustrated  in  Fig.  31 ;  in  vacuum 
systems  it  is  well  to  have  a  gate  valve  in  the  air  line. 

Offsets  in  steam  and  return  piping  should  preferably  be  made  with 
90-deg  ells  but  occasionally  fittings  of  other  angles  are  used,  and  in  such 
cases  the  length  of  the  diagonal  offset  will  be  found  as  shown  in  Fig.  32. 

Branches  from  steam  mains  in  one-pipe  gravity  steam  systems  should 
use  the  preferred  connection  shown  in  Fig.  33,  but  where  radiator  condensa- 
tion does  not  flow  back  into  the  main  the  acceptable  method  shown  in  the 
same  figure  may  be  used.  This  acceptable  method  has  the  advantage^ 
giving  a  perfect  swing  joint  when  connected  to  the  vertical  riser  or  radia- 
tor connection,  whereas  the  preferred  connection  does  not  give  this  swing 
without  distorting  the  angle  of  the  pipe.  Runouts  from  the  steam  main 
are  usually  made  about  5  ft  long  to  provide  flexibility  for  movement  in 
the  main. 

Dirt  pockets,  desirable  on  all  systems  employing  thermostatic  traps, 
should  be  so  located  as  to  protect  the  traps  from  scale  and  muck  which 
will  interfere  with  their  operation.  Dirt  pockets  are  usually  made  8  in. 
to  12  in.  deep  and  serve  as  receivers  for  foreign  matter  which  otherwise 
would  be  carried  into  the  trap.  They  are  constructed  as  shown  in  Fig.  34. 

On  vapor  systems  where  the  end  of  the  steam  main  is  dripped  down 
into  the  wet  return,  the  air  venting  at  the  end  of  the  main  is  accomplished 
by  an  air  vent  passing  through  a  thermostatic  trap  into  the  dry  return 
line  as  shown  in  Fig.  35.  On  vacuum  systems  the  ends  of  the  steam  mains 
are  dripped  and  vented  into  the  return  through  thermostatic  drip  traps 
opening  into  the  return  line.  The  same  method  may  be  used  in  atmos- 
pheric systems.  The  cooling  leg  (Fig.  36)  is  for  cooling  the  condensation 

555 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

sufficiently  before  it  reaches  the  trap  so  the  trap  will  not  be  held  shut  by 
too  high  a  temperature.  On  down-feed  systems  of  atmospheric,  vapor, 
and  vacuum  types,  the  bottom  of  the  steam  risers  are  dripped  in  the 
manner  shown  in  Fig.  37. 


PROBLEMS  l\  PRACTICE 

1  •  What  is  the  equivalent  length  of  run  in  a  steam  system? 

It  is  the  length  of  straight  pipe  which  will  have  the  same  friction  and  pressure  drop  as  a 
shorter  length  of  pipe  of  the  same  size  with  accompanying  valves,  tees,  elbows,  and  other 
fittings  will  have,  when  both  pipes  are  carrying  the  same  amount  of  steam  at  the  same 
pressure. 

2  •  When  the  size  of  pipe  is  still  undetermined,  what  arbitrary  percentage  is 
usually  added  to  the  actual  length  to  obtain  the  equivalent  length? 

Usually  100  per  cent;  in  other  words,  the  actual  length  is  doubled  to  allow  for  the  added 
drop  produced  by  the  valves,  tees,  elbows,  and  other  fittings. 

3  •  What  are  the  major  factors  to  be  considered  in  determining  the  flow  of 
steam  in  pipes? 

a.  The  initial  steam  pressure  available  and  the  total  pressure  drop  allowable  between  the 
source  of  steam  supply  and  the  end  of  the  return  system.    The  pressure  drop  should 
never  exceed  one  half  of  the  initial  pressure. 

b.  The  maximum  steam  velocity  allowable.     When  condensate  is  flowing  against  the 
steam,  the  velocity  must  not  be  so  great  as  to  produce  water  hammer,  or  hold  up 
water  in  parts  of  the  system  until  the  steam  flow  is  reduced  sufficiently  to  permit  the 
water  to  pass.    The  velocity  at  which  disturbances  take  place  depends  upon : 

1.  Size  of  pipe. 

2.  Whether  pipe  is  vertical  or  horizontal. 

3.  Pitch  or  grade  of  pipe. 

4.  Quantity  of  water  flowing  against  steam. 

c.  The  equivalent  length  of  run  from  the  source  of  steam  supply  to  the  farthest  heating 
unit,  with  allowance  for  friction  in  pipe  fittings  and  valves. 

4  •  Name  three  fundamental  considerations  in  designing  the  piping  system 
for  steam  heating. 

a.  Provision  for  the  distribution  of  suitable  quantities  of  steam  to  the  various  heating 
units. 

b.  Provision  for  the  return  of  condensate  from  the  radiators  and  piping  to  the  boiler. 

c.  Provision  of  means  for  expelling  air  from  the  radiators  and  piping. 

5  •  Why  is  the  proper  reaming  of  the  ends  of  pipe  necessary? 

The  capacities  of  pipes  depend  upon  the  free  area  available  for  flow.  In  cutting  the  pipe 
this  area  may  be  restricted  by  a  burr,  which  may  decrease  the  capacity  of  a  pipe  more 
than  25  per  cent  in  the  smaller  pipe  sizes. 

6  •  a.  What  are  the  major  factors  to  he  considered  when  selecting  a  pressure 

reducing  valve? 
b.  How  should  such  valve  be  installed? 

a.  The  initial  pressure  of  the  steam  must  be  considered  along  with  the  desired  reduced 
pressure.    The  connected  load  to  be  supplied  must  be  known  in  square  feet  of  equiva- 

556 


CHAPTER  32 — PIPING  FOR  STEAM  HEATING  SYSTEMS 


lent  direct  radiation  or  in  pounds  of  steam  per  hour.  For  operation  with  a  continuous 
load,  a  semi-balanced  or  double  seated  valve  operated  by  a  diaphragm  gives  good 
results.  Where  the  load  is  intermittent,  as  in  process  work  or  with  thermostatically 
controlled  blast  heaters,  a  so-called  dead  end  or  single  seated  valve  should  be  used. 

The  pressure  reducing  valve  should  be  installed  in  a  horizontal  line  with  a  gate  valve 
on  each  side,  and  with  a  by-pass  operated  by  a  valve.  The  pressure  balancing  pipe 
from  the  diaphragm  chamber  should  be  connected  into  the  top  or  side  of  the  low 
pressure  main  not  less  than  15  ft  from  the  reducing  valve. 

7  •  What  is  the  usual  expansion  allowance  and  how  it  is  compensated  for  in 
heating  system  supply  risers? 

The  expansion  of  low  pressure  stearn  piping  is  normally  taken  as  1J^  to  1^  in.  per  100 
ft  of  pipe.  With  a  five  story  building  a  double  swing  connection  between  the  riser  and  the 
main  will  suffice.  In  buildings  between  5  and  10  stories  high  the  riser  should  be  anchored 
near  its  center  and  have  double  swing  connections  to  the  main.  For  taller  buildings 
expansion  loops  or  riser  offsets  are  used  which  are  capable  of  handling  a  length  of  riser 
reaching  5  stories  in  either  direction  from  the  joint.  The  risers  are  anchored  at  each 
alternate  5  stories.  All  radiators  must  have  double  swing  connections,  and  those  con- 
nected above  where  the  riser  is  anchored  must  be  given  greater  pitch  to  insure  their 
having  proper  grade  when  the  riser  is  heated. 

8  •  Why  should  all  boiler  steam  supply  tappings  be  used  full  size? 

In  order  to  operate  at  low  steam  velocities  so  the  water  in  suspension  can  separate  from 
the  steam  and  remain  in  the  boiler. 

9  •  What  is  the  Underwriters  Loop  or  the  Hartford  Connection? 

An  arrangement  of  piping  on  the  returns  to  low  pressure  boilers  wherein  the  return  line 
is  raised  up  nearly  to  the  water  line  of  the  boiler  and  is  then  dropped  back  and  con- 
nected to  the  boiler  return  inlet;  the  high  point  is  connected  by  a  balance  pipe  to  the 
steam  runout  from  the  boiler  on  the  boiler  side  of  all  stop  valves.  With  this  loop  no 
check  valve  is  required,  and  water  cannot  be  backed  out  of  the  boiler  and  into  the  return 
at  a  point  lower  than  the  invert  of  the  pipe  at  the  top  of  the  loop. 

10  •  What  are  the  important  factors  in  making  radiator  connections? 

Connections  to  radiators  should  be  made  as  direct  as  possible,  of  proper  size,  with  ample 
pitch  of  piping  and  allowance  for  expansion. 

11  •  Why  should  careful  attention  be  given  to  proper  dripping  and  drainage 
of  steam  piping? 

The  steam  mains  and  risers  must  be  quickly  drained  of  condensate  and  where  necessary 
vented  of  air  in  order  to  obtain  a  sufficient  supply  of  steam  to  the  radiators.  Proper 
drainage  is  also  necessary  to  insure  a  noiseless  heating  system. 

12  •  What  is  the  limit  of  pressure  drop  usually  recommended  in  a  vacuum 
system? 

Not  over  y%  Ib  (2  02)  per  100  ft  of  equivalent  run,  and  not  over  1  Ib  total  drop. 

13  •  When  steam  and  condensation  are  flowing  in  the  same  direction,  what  is 
the  maximum  total  pressure  drop  which  should  be  used? 

The  maximum  total  pressure  drop  should  not  exceed  one  half  of  the  initial  steam  pressure. 

14  •  What  does  a  proper  installation  of  a  pressure  reducing  valve  include? 

A  strainer  in  front  of  the  pressure  reducing  valve;  a  gate  valve  in  front  of  the  strainer;  a 
gate  valve  after  the  reducing  valve;  a  by-pass  around  the  two  gate  valves,  strainer f  and 
pressure  reducing  valve;  and  a  globe  valve  in  the  by-pass.  Sometimes  a  safety  valve  on 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

the  low  pressure  side  and  pressure  gages  on  both  sides  are  installed.  The  high  pressure 
line  should  be  dripped  just  before  the  high  pressure  steam  enters  the  pressure  reducing 
valve  assembly, 

15  •  Will  a  pressure  reducing  valve  which  is  reducing  the  steam  pressure  from 
100  Ib  gage  to  50  Ib  gage  pass  more  or  less  steam  than  the  same  valve  when 
reducing  the  steam  pressure  from  100  Ib  gage  to  5  Ib  gage? 

The  valve  will  pass  practically  the  same  volume  of  steam  in  each  case  as  the  velocity  of 
steam  flowing  through  an  orifice  shows  no  material  increase  after  the  reduced  absolute 
pressure  has  fallen  to  58  per  cent  of  the  initial  absolute  pressure.  Because  of  its  greater 
density,  the  weight  of  steam  passed  will  be  greater  in  the  case  of  the  reduction  to  50  Ib 
gage. 


558 


Chapter  33 

HOT  WATER  HEATING  SYSTEMS 
AND  PIPING 

One-  and  Two-Pipe  Systems,  Selecting  Pipe  Sizes,  Forced  Circu- 
lation, Effect  of  Variations  in  Pipe  Sizes,  Gravity  Circulation, 
Mechanical  Circulation,  Expansion  Tanks,  Installation  Details 

A  HOT  water  heating  system  is  one  in  which  water  is  the  medium  by 
which  heat  is  carried  through  pipes  from  the  boiler  to  the  heating 
units.  There  are  two  general  types,  namely,  forced  circulation  and  gravity 
circulation  systems.  In  the  former  the  pressure  head  maintaining  flow  is 
produced  mechanically,  whereas  in  the  latter  the  pressure  head  is  pro- 
duced by  the  differences  in  weight  of  the  water  in  the  flow  and  in  the 
return  risers. 

The  fundamental  rule  in  the  design  of  a  hot  water  system  is  that  the 
total  friction  and  resistance  head  in  any  circuit  must  equal  the  pressure 
head  causing  the  water  to  flow  in  the  same  circuit. 

In  designing  a  hot  water  heating  system,  it  is  necessary  to  determine : 

1.  The  heat  losses  of  the  rooms  or  spaces  to  be  heated.    (See  Chapter  7.) 

2.  The  size  and  type  of  boiler.     (See  Chapter  25.) 

3.  The  location,  type,  and  size  of  heating  units.     (See  Chapter  30.) 

4.  The  method  of  piping. 

5.  Suitable  pipe  sizes. 

6.  The  type  and  size  of  circulating  pump  (if  forced  circulation). 

7.  The  type  and  size  of  expansion  tank. 

The  unit,  a  square  foot  of  equivalent  direct  radiation,  E,DR,  has  been  used 
for  many  years  for  rating  purposes  in  both  steam  and  hot  water  systems,  but 
its  use,  especially  in  hot  water  systems,  has  always  resulted  in  complications 
and  confusion.  It  is  the  plan  of  THE  GUIDE  to  eventually  eliminate  this 
empirical  expression  and  to  substitute  a  logical  unit  based  on  the  Btu.  The 
Mb,  the  equivalent  of  1000  Btu,  and  the  Mbh,  the  equivalent  of  1000  Btu 
per  hour,  which  have  been  approved  by  the  A.S.H.V.E.,  are  used  in  this 
chapter  on  hot  water  systems  to  replace  the  square  foot  of  radiation  formerly 
used. 

ONE-  AND  TWO-PIPE  SYSTEMS 

Pipe  systems  may  be  divided  into  two  general  types,  namely,  two-pipe 
and  one-pipe  systems.  In  a  two-pipe  system  the  piping  is  arranged  so  that 
the  water  flows  through  only  one  radiator  during  a  circuit  through  the 
system,  so  that  all  radiators  are  supplied  with  water  at  practically  the 
same  temperature  as  that  in  the  boiler.  In  a  one-pipe  system,  the  water 
flows  through  more  than  one  radiator  during  its  circuit.  In  that  case,  the 

559 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

first  radiator  receives  the  hottest  water;  the  second  radiator,  somewhat 
cooler  water;  the  third  one,  still  cooler;  and  so  on.  As  the  temperature  of 
the  water  supplied  to  a  radiator  is  lowered,  the  size  of  the  radiator  must 
be  increased  and,  consequently,  the  total  heating  surface  for  a  one-pipe 
system  must  be  greater  than  that  for  a  two-pipe  system  for  the  same 
service. 

Two-pipe  systems  may  be  divided  into  two  classes,  direct  return  sys- 
tems (Fig.  1),  and  reversed  return  systems  (Fig.  2).  In  a  direct  return 
system  the  water  returns  to  the  heater  by  a  direct  route  after  it  has 
passed  through  its  radiator  and,  as  a  result,  the  paths  through  the  three 
radiators  shown  in  Fig.  1  are  of  unequal  lengths,  the  path  through  the 
first  radiator  being  the  shortest  and  that  through  the  third  radiator,  the 
longest.  In  a  reversed  return  system,  the  water  returns  to  the  heater  by 
an  indirect  route  after  it  has  passed  through  the  radiators,  so  that  the 
paths  leading  through  the  three  radiators  shown  in  Fig.  2  are  practi- 
cally of  equal  length. 

The  reversed  return  system  has  an  advantage  over  the  direct  return 
system  in  that  it  is  more  likely  to  function  satisfactorily  even  though  the 


FIG.  1.    A  DIRECT  RETURN  SYSTEM 


FIG.  2.    A  REVERSED  RETURN  SYSTEM 


pipe  system  is  not  accurately  designed.  For  example,  if  in  Fig.  2  all  pipes 
are  of  one  size,  each  of  the  three  radiators  will  receive  approximately  the 
same  quantity  of  hot  water  because  the  three  paths  are  practically  of 
-equal  length,  whereas  in  Fig.  1,  if  all  pipes  are  of  the  same  size,  Radiator 
1  will  receive  more  water  than  the  others  because  the  path  through  it  is 
shorter  than  those  through  the  other  radiators.  As  a  result,  Radiator  1 
will  be  filled  with  water  at  a  higher  average  temperature  than  the  re- 
maining two  radiators,  and  will  therefore  dissipate  more  heat.  To  pre- 
vent this  unequal  distribution  of  heat  it  is  necessary  to  throttle  the  paths 
through  Radiators  1  and  2  so  that  the  friction  heads  of  the  three  paths  are 
<equal  when  each  radiator  receives  its  proper  quantity  of  water. 

A  comparison  of  Fig.  1  and  Fig.  2  may  suggest  that  a  reversed  return 
system  requires  considerably  longer  mains  than  a  direct  return  system. 
This  is  not  always  the  case.  For  example,  note  the  reversed  return 
system  of  Fig.  3. 

PIPE  SIZES 

The  pressure  heads  available  in  forced  circulation  systems  are  much 
larger  than  those  in  gravity  circulation  systems,  consequently,  higher 
velocities  may  be  used  in  designing  the  system,  with  the  result  that  smaller 
pipes  may  be  selected  and  the  first  cost  of  the  installation  reduced.  As 
the  pipes  of  a  heating  system  are  reduced  in  size,  the  necessary  increase  in 

560 


CHAPTER  33 — HOT  WATER  HEATING  SYSTEMS  AND  PIPING 

the  velocity  of  the  water  increases  the  cost  of  operating  the  circulating 
pump.  There  is  an  optimum  velocity  of  the  water  in  a  heating  system  for 
which  the  sum  of  the  cost  of  the  system  and  the  cost  of  its  operation  is  a 
minimum.  This  velocity  should  be  determined  by  calculation  for  the- 
particular  system  under  consideration. 

Since  the  velocities  in  forced  circulation  systems  are  higher  than  those 
in  gravity  circulation  systems,  and  since  the  friction  heads  in  a  heating 
system  vary  almost  as  the  squares  of  the  velocities,  a  given  error  in  the 
calculation  or  assumption  of  a  velocity  is  less  important  in  a  forced  circu- 
lation system  than  in  a  gravity  circulation  system  and,  consequently,  it 


t/z" 


I'A." 


54 


C|  14-2. 


L. 
G 


12 


_r_ 

H 


GG 


4- 


GG 


54- 


ilk" 


At 


FIG.  3.    A  FORCED  CIRCULATION  REVERSED  RETURN  SYSTEM* 

aThis  system  could  be  divided  into  two  branches.  This  would  permit  the  use  of  smaller  pipes  and  would 
produce  only  slight  changes  in  the  total  length  of  the  pipe.  It  is  shown  as  a  single  system  here  simply  to- 
illustrate  the  method  of  determining  pipe  sizes  by  means  of  pipe  size  tables.  Note  that  the  numbers  on 
the  radiators  indicate  thousands  of  Btu  per  hour  (Mbh)  and  not  square  feet. 

is  easier  to  design  a  satisfactory  forced  circulation  system  than  a  satis- 
factory gravity  circulation  system. 


FORCED  CIRCULATION 

The  following  examples  will  illustrate  the  procedure  to  be  followed  in 
designing  forced  circulation  systems: 

Example  1.  Assume  that  the  paths  through  the  five  radiators  shown  in.  Fig.  3  consist 
each  of  150  ft  of  mains,  5  ft  of  radiator  connections,  1  boiler,  1  radiator,  1  radiator  valve, 
10  ells,  and  2  tees.  Design  the  piping  for  this  system. 

Solution.  The  friction  heads  in  the  boiler,  radiator,  valve,  and  tee  may  be  expressed 
in  terms  of  the  friction  head  in  one  elbow  according  to  the  values  given  in  Table  1. 
Having  done  this,  each  of  the  five  circuits  is  taken  as  155  ft  of  pipe  and  about  24  elbow 
equivalents.  The  friction  head  of  one  elbow  is  approximately  equivalent  to  that  in  a 
pipe  having  a  length  equal  to  25  diameters.  Assuming  that  the  average  pipe  size  in  this 
case  will  be  about  1  l/i  in.,  one  elbow  equivalent  may  be  placed  equal  to  about  3  ft  of  pipe 
and  the  total  length  of  the  circuit  equivalent  to  about  227  ft  of  pipe. 

Having  determined  the  equivalent  pipe  length,  assume  the  rate  at  which  the  water  i* 

561 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


1000 


j 


'SO  .0<  500  \W 

HCAT    CONVEYELD      PE*     HOUB     IN       1000       5,  T.  U, 

Tfcc  DirrcccNCC  JN  TCMPCPATUEC  or    THE  WATER   IN   Tnt.  FLOW  AND  SCTURN  Itotw 


10000 


FIG,  4.   FRICTION  HEADS  IN  PIPES  FOR  A  20  F  TEMPERATURE  DIFFERENCE 
OF  THE  WATER  IN  THE  FLOW  AND  RETURN  LINES 


CHAPTER  33 — HOT  WATER  HEATING  SYSTEMS  AND  PIPING 

to  be  forced  through  the  system.  This  rate  may  vary  widely.  The  water  may  flow 
through  the  radiator  so  that  it  will  cool  10  deg  or  20  deg  or  any  other  reasonable  number 
of  degrees.  In  this  case,  assume  a  10-deg  drop.  Since  the  system  is  to  dissipate  66,000 
Btu  per  hour  (66  Mbh),  the  pump  must  circulate  6600  Ib  of  water  per  hour  or  13.8  gpm 
based  on  the  actual  density  of  water  of  7.99  Ib  per  gallon  at  215  F.  One  gallon  of  water 
per  minute  at  this  density  will  deliver  9600  Btu  per  hour  (9.6  Mbh)  with  a  temperature 
drop  of  20  F. 

TABLE  1.     ELBOW  EQUIVALENTS'* 

1  90-deg  elbow._ ., 1.0 

1  45-deg  elbow 0.7 

1  90-deg  long  turn  elbow 0.5 

1  open  return  bend 1.0 

1  open  gate  valve 0.5 

1  open  globe  valve 12.0 

1  angle  radiator  valve 2.0 

1  radiator. 3.0 

1  heater.-.. 3.0 

1  tee (Noteb) 


aThe  loss  of  head  in  one  elbow  can  be  expressed  in  terms  of  the  velocity  head  by  the  formula: 

1/2 


where 


h  =  the  loss  of  head  in  feet,  v  =*  the  velocity  of  approach  in  feet  per  second, 
and  2g  =  64.4  ft  per  second  per  second. 


l>The  loss  of  head  in  tees  when  water  is  diverted  at  right  angles  through  a  branch  of  the  tee  varies  with 
the  per  cent  diverted.  When  the  water  diverted  is  less  than  60  per  cent  of  that  approaching  the  tee,  the 
loss  of  head,  in  elbow  equivalents,  may  be  expressed  as  follows: 

**  -  3 

where 

A«  «=  the  loss  of  head  in  elbow  equivalents,  vi  =  the  velocity  of  approach, 
vt  =    the  velocity  of  water  diverted  at  right  angles. 

Values  in  elbow  equivalents  for  the  most  common  percentages  of  water  diverted  in  a  Ixlxl-in.  tee  are 
as  follows: 

25%  ______  ...............................................................................  16.0 

33%  ........................................................................................    9.0 

100^r////rr//r////™^  1.8 

For  other  percentages  the  approximate  values  may  be  secured  by  interpolation.  When  the  water  is 
diverted  from  the  tee  into  a  smaller  size  branch,  as  in  a  IxlxK-ia-  tee,  approximate  values  may  be  secured 
by  means  of  Formula  2. 

The  next  step  in  the  design  is  to  assume  the  velocity  at  which  the  water  is  to  circulate 
through  the  system.  This  also  may  vary  materially.  As  the  velocity  is  increased,  the 
sizes  of  the  pipes  and  the  cost  of  the  system  are  decreased,  but  the  cost  of  operating  the 
circulating  pump  is  increased.  The  designing  engineer  should  make  a  careful  study  to 
determine  the  velocity  which  will  produce  the  most  economical  installation  for  the 
particular  case  in  hand.  In  this  case,  assume  a  velocity  of  about  1H  fps  for  a  lJ<C-m. 
pipe. 

Reference  to  Fig.  4  shows  that  for  a  IJ^-in.  pipe  and  a  velocity  of  18  in.  per  second, 
the  friction  head  is  about  100  milinches  per  foot,  or  about  2  ft  for  a  circuit  of  227  ft,  if 
the  pipe  sizes  for  that  circuit  are  chosen  so  that  the  average  friction  head  is  about 
100  milinches  per  foot  of  pipe. 

The  pipe  sizes  may  now  be  selected  from  Fig.  4  by  making  allowance  for  the  fact  that 
Fig.  4  is  based  on  a  temperature  drop  of  20  F  and  that  the  system  to  be  designed  is  to 
have  a  temperature  drop  of  only  10  F  as  follows:  Sections  AB  and  KA  carry  66,000 
Btu  per  hour  (66  Mbh)  with  a  temperature  drop  of  10  F;  if  the  temperature  drop  were 
20  F  these  sections  would,  with  the  same  velocity  and  the  same  friction  head,  carry 
132,000  Btu  per  hour  (132  Mbh).  Hence,  refer  to  Fig.  4  for  132,000  Btu  and  a  unit 
friction  head  of  100  milinches,  and  note  that  the  correct  size  would  be  about  halfway 
between  a  l%~in,  and  a  2-in.  pipe.  Therefore,  select  a  !J4-in.  pipe  for  Section  AB  and 

563 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


a  2-in.  pipe  for  Section  KA  .  The  pipe  sizes  for  the  remaining  eight  sections  and  for  the 
radiator  connections  can  be  selected  in  the  same  manner  and  recorded  on  the  pipe 
diagram  as  shown. 

The  circulating  pump  for  the  system  should  be  one  which  has  its  highest  efficiency 
when  it  is  delivering  13.8  gpm  against  a  head  of  2  ft, 

If  a  number  of  heating  systems  are  to  be  designed  for  similar  conditions, 
i.e.,  for  a  total  friction  head  of  2  ft  and  a  temperature  drop  through  the 
radiators  of  10  F  when  the  maximum  quantity  of  heat  is  being  delivered 
to  the  building,  a  table  such  as  Table  2  may  be  prepared  from  the  data  of 
Fig.  4.  Having  this  table,  the  pipe  sizes  for  the  system  of  Example  1  can 
be  easily  selected.  For  example,  for  Sections  BC  and  JK,  each  supplying 
54  Mbh,  the  equivalent  pipe  length  of  the  system  is  227  ft.  In  the  table 
the  length  shown  nearest  to  this  length  is  200  ft.  In  the  200-ft  column, 
a  1^-in.  pipe  is  slightly  too  small  and  a  2-in.  pipe  is  too  large.  The 
l3/£-in.  pipe  will  therefore  be  selected.  For  Sections  CD  and  IJ,  supplying 
42  Mbh,  a  IJ^-in.  pipe  is  too  small  and  a  1^-in.  pipe  is  too  large,  so  1J^  in- 
will  be  selected  for  the  flow  and  1J^  in.  for  the  return  line.  For  larger 


TABLE  2.    CAPACITIES  OF  PIPES  IN  Mbh  (1000  Bxu  PER  HOUR)  AND  Velocities  of 

Water  in  Pipes  in  Inches  per  Second  FOR  FORCED  CIRCULATION  SYSTEMS 

WITH  A  TOTAL  FRICTION  HEAD  OF  2  FT  AND  FOR  A  MAXIMUM 

TEMPERATURE  DROP  OF  10  Fa 


I 

2 

3 

4 

5 

6 

7 

8 

9 

PIPE 
SIZE 
(INCHES) 

EQUIVALENT 
LENGTH 
OP  PIPE 

(FEBTb) 

EQUIVALENT  TOTAL  LENGTH  OF  PIPE  IN  FEET  IN  LONGEST  CIRCUIT 

100 

150 

200 

250 

300 

350 

400 

UNIT  FRICTION  HEAD,  IN  MILINCHES 

240 

160 

120 

96 

80 

69 

60 

1A 

1 

8.8 
15 

4.8 
12 

4.1 
10 

8.4 

9 

2.9 
8 

2.6 
7.5 

8.4 

7 

% 

2 

18.8 
18 

10.  8 
14 

8.6 
12 

7.3 
11 

6.8 
10 

6.0 
9 

5.5 
8,5 

i 

2.3 

86.0 
22 

19.8 
17 

16.3 
15 

14-4 
13 

12.5 
12 

12.0 
11 

11.1 
10.5 

IK 

3.0 

52.8 
27 

40.8 
21 

84.8 
18 

SI.  8 
16 

27.8 
15 

26.4 
14 

84-0 
13 

iH 

3.5 

79.2 
30 

60.7 
23 

51.2 
20 

45.6 
18 

40.8 
16 

40.0 
15 

86.0 
14 

2 

4.0 

158.8 
36 

120.0 
28 

104.0 
24 

93.5 
22 

86.4 
20 

81.5 
18 

78.8 
17 

2M 

6.0 

250.  0 
41 

198.0 
32 

164.5 
28 

149.0 
25 

189.2 
22 

185.8 
21 

122.5 
19 

3 

6.5 

444-0 
48 

348.0 
37 

294.0 
32 

270.0 
29 

254.0 
26 

840.0 
24 

228.0 
22 

«For  other  temperature  drops  the  capacities  of  pipes  are  to  be  changed  correspondingly.  For  example, 
for  a  temperature  drop  of  30  F,  the  capacities  shown  in  this  table  are  to  be  multiplied  by  3.  The  velocities 
remain  unchanged, 

^Approximate  length  of  pipe  in  feet  equivalent  to  one  elbow  in  friction  head.  This  value  varies  with  the 
velocity. 

564 


CHAPTER  33 — HOT  WATER  HEATING  SYSTEMS  AND  PIPING 


systems,  it  will  be  economical  to  operate  with  higher  friction  heads,  and 
tables  may  be  prepared  similar  to  Tables  3  and  4,  which  are  based  on 
total  friction  heads  of  6  and  18  ft,  respectively. 

Example  2,  Design  a  direct  return  two-pipe  forced  circulation  system  for  the  layout 
shown  in  Fig.  5.  For  this  system  the  length  of  the  pipe  line  from  the  boiler  to  the 
highest  radiator  on  the  farthest  riser  and  back  to  the  boiler  is  about  250  ft.  There  are 
about  16  elbow  equivalents  having  an  equivalent  pipe  length  of  about  50  ft,  so  the  total 
equivalent  pipe  length  is  about  300  ft. 

Solution.  The  same  pipe  size  tables  may  be  used  as  those  developed  for  the  reversed 
return  system  of  Fig.  3.  Since  this  system  is  somewhat  larger  than  that  shown  in  Fig.  3, 
Table  3  which  provides  for  a  friction  head  of  6  ft  may  be  used  instead  of  Table  2  which 
provides  for  a  friction  head  of  only  2  ft. 

Referring  to  the  column  for  an  equivalent  total  length  of  300  ft  for  Sections  AB  and 
KA,  each  supplying  117.6  Mbh,  it  will  be  found  that  a  1  J^-in.  pipe  is  too  small  and  a  2-in. 
pipe  is  too  large.  Consequently,  a  lj^-in.  pipe  is  selected  for  the  flow  line  AB,  and  a 
2-in.  pipe  for  the  return  line  KA.  For  Sections  BC  and  JK,  each  supplying  88  Mbh,  a 
IJ^-in.  pipe  is  only  slightly  too  small  and  it  is  selected.  The  remaining  pipe  sizes  are 
selected  in  a  similar  manner  and  recorded  in  Fig.  5.  For  a  temperature  drop  of  10  F, 
24.5  gpm  of  water  must  be  circulated.  The  pump  to  select  is  one  which  has  its  highest 
efficiency  when  it  is  delivering  24.5  gpm  against  a  6-ft  head. 

TABLE  3.     CAPACITIES  OF  PIPES  IN  Mbh  (1000  Bxu  PER  HOUR)  AND  Velocities  of 

Water  in  Pipes  in  Inches  per  Second  FOR  FORCED  CIRCULATION  SYSTEMS 

WITH  A  TOTAL  FRICTION  HEAD  OF  6  FT  AND  FOR  A  MAXIMUM 

TEMPERATURE  DROP  OF  10  Fa 


1 

2 

3 

4 

5 

6 

1 

8 

PIPE 
SIZE 
(INCHES) 

EQUIVALENT 
LENGTH 
OF  PIPE 

(FffiETb) 

EQUIVALENT  TOTAL  LENGTH  OF  PIPE  IN  FEET  IN  LONGEST  CIRCUIT 

200 

300 

400 

600 

800 

1000 

UNIT  FRICTION  HEAD,  IN  MILINCHES 

360 

240 

180 

120 

90 

72 

1A 

1 

7.4 
18 

6.0 

15 

5.0 
13 

8.8 
10 

a  S.4 
9 

3.1 
7.5 

U 

2 

15.8 
22 

12.7 
18 

10.8 
16 

8.4 
12 

7.7 
11 

6.7 
9 

i 

2.5 

30.0 
27 

24.0 
22 

20.4 
19 

15.8 
15 

13.9 
13 

18.5 
11 

IK 

3.3 

64-8 
33 

6Q.5 
26 

44*4 
23 

33.6 
18 

30.0 
16 

26.8 
14 

i« 

4.0 

96.0 
37 

76.8 
31 

64.8 
26 

50.1 
20 

44.  7 
18 

40.  8 
15 

2 

5.0 

19%.  0 
44 

15S.O 
36 

1SO.O 
30 

100.1 
24 

90.0 
21 

78.0 
18 

VA 

6.0 

$00.0 
50 

844-0 
41 

206.0 

35 

161.0 
26 

144.0 
24 

130.0 
21 

3 

7.5 

550.0 
58 

436.  0 
48 

S68.0 
42 

287.0 
32 

249.0 
27 

228,0 
24 

aFor  other  temperature  drops  the  capacities  of  pipes  are  to  be  changed  correspondingly.  For  example, 
for  a  temperature  drop  of  30  F,  the  capacities  shown  in  this  table  are  to  be  multiplied  by  3.  The  velocities 
remain  unchanged. 

^Approximate  length  of  pipe  in  feet  equivalent  to  one  elbow  in  friction  head.  This  value  varies  with 
the  velocity. 

565 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


To  secure  a  correct  distribution  of  hot  water  among  the  several  risers  it  is  necessary, 
as  previously  stated,  to  introduce  special  resistances  to  balance  the  several  risers,  as 
follows: 

The  first  riser  is  80  ft  nearer  the  boiler  than  the  fifth  riser.  In  order  that  the  two  may 
be  balanced,  i.e.,  that  they  may  operate  under  equal  pressure  heads,  resistance  must  be 
added  to  the  first  riser  equal  to  the  friction  head  in  the  80  ft  of  flow  main  from  B  to  F 
plus  that  in  the  80  ft  of  return  main  from  G  to  K. 

It  will  be  noted  from  Table  3  that  the  unit  friction  head  is  about  240  milinches  per 
foot.  The  total  friction  head  in  the  flow  and  return  mains  between  the  first  and  fifth 
risers  is  therefore  160  X  240  or  38,400  milinches,  or  a  little  more  than  3  ft,  which  must  be 
supplied  by  additional  resistance  in  the  first  riser  to  prevent  its  having  an  advantage 
over  the  fifth  riser. 

This  resistance  can  be  supplied  by  a  calibrated  and  adjusted  modulating  valve  or  by 
an  orifice  resistor  in  a  union.  If  the  orifice  resistor  is  to  be  used,  its  size  may  be  selected 
from  Table  5  as  follows: 

The  lower  part  of  the  first  flow  riser  supplies  28.8  Mbh.  According  to  Table  3,  it 
should  be  a  1-in.  pipe  and  would  have  a  velocity  of  22  in.  per  second,  if  it  were  supplying 
24  Mbh.  Since  it  is  supplying  28.8  Mbh,  the  velocity  will  be  about  26  in.  per  second. 
From  Table  5  it  will  be  found  that  for  a  1-in.  pipe  and  a  velocity  of  24  in.  per  second,  an 
0.45-in.  orifice  will  produce  a  loss  of  head  of  37,000  milinches.  For  a  velocity  of  26  in. 
per  second,  the  loss  of  head  will  be  somewhat  more,  probably  about  43,000  milinches;  the 

TABLE  4.    CAPACITIES  OF  PIPES  IN  Mbh  (1000  BTU  PER  HOUR)  AND  Velocities  of 

Water  in  Pipes  in  Inches  per  Second  FOR  FORCED  CIRCULATION  SYSTEMS 

WITH  A  TOTAL  FRICTION  HEAD  OF  18  FT  AND  FOR  A  MAXIMUM 

TEMPERATURE  DROP  OF  10  Fa 


1 

2 

3 

4 

s 

6 

7 

PIPE 
SIZE 
(INCHES) 

EQTTIVALHNT 
LENGTH 

OF  PIPE 

(Firob) 

EQUIVALENT  TOTAL  LENGTH  OP  PIPE  IN  FEET  IN  LONGEST  CIRCUIT 

200 

400 

600 

800 

1000 

UNIT  FRICTION  HuAii,  IN  MILINCHES 

1080 

540 

360 

270 

216 

1A 

1.0 

IS.  7 
32 

8.6 

23 

7.8 
18 

6.2 
15 

5.6 
13 

H 

2.0 

87.5 
40 

18.7 
28 

15.1 
22 

18.7 
19 

11.5 
17 

1 

2.5 

55.0 
4$ 

86.8 
34 

so.o 

27 

86.4 
23 

88.6 
20 

IK 

3.0 

188.  0 
59 

81.5 
42 

66.0 
33 

58.S 
28 

50.5 
25 

1H 

4.0 

188.0 
66 

188.0 

46 

98.8 

37 

86.8 
31 

74.8 
27 

2 

5.0 

371,0 
80 

£58.0 
56 

201.0 
45 

180.0 
38 

151.0 
33 

2^ 

7.0 

598.0 
91 

407.0 
65 

S2S.O 
51 

887.0 
43 

840.0 
3$ 

3 

9.0 

1110.  0 
107 

790.0 
76 

598.0 
60 

587.0 
51 

443.0 
44 

*For  other  temperature  drops  the  capacities  of  pipes  are  to  be  changed  correspondingly.  For  example, 
for  a  temperature  drop  of  30  F,  the  capacities  shown  in  this  table  are  to  be  multiplied  by  3.  The  velocities 
remain  unchanged. 

^Approximate  length  of  pipe  in  feet  equivalent  to  one  elbow  in  friction  head.  This  value  varies  with 
the  velocity. 

566 


CHAPTER  33 — HOT  WATER  HEATING  SYSTEMS  AND  PIPING 

difference  between  it  and  the  required  resistance  will  be  about  10  per  cent,  which  is  per- 
missible, and  the  0.45-in.  orifice  is  selected. 

The  sizes  of  the  orifice  resistors  for  the  second,  third,  and  fourth  risers  are  selected  in 
a  similar  manner  and  found  to  be  0.45  in,,  0.50  in.,  and  0.55  in.,  respectively. 

If  the  design  of  the  system  of  Fig.  5  is  to  be  extremely  refined,  the 
gravity  pressure  heads  produced  by  the  risers  should  be  taken  into  con- 
sideration. With  water  at  220  F  and  210  F,  respectively,  in  the  risers,  the 
gravity  head  is  50  milinches  per  foot  of  water  column  or  25  milinches  per 
foot  of  flow  and  return  pipe.  The  pump  pressure  head  in  this  case  is  240 
milinches  per  foot  of  pipe,  and  the  gravity  head,  being  only  one  tenth  as 
large  as  the  pump  head,  may  be  neglected  without  serious  error.  This  is 
generally  done. 

Temperatures  of  220  F  and  210  F  would  be  used  only  during  the  coldest 


FIG,  5.    A  FORCED  CIRCULATION  DIRECT  RETURN  SYSTEM 

weather  for  which  the  system  is  designed.  At  other  times  the  tempera- 
tures would  be  lower,  the  temperature  drop  smaller,  and  the  gravity  heads 
smaller.  The  pump  pressure  head  remains  constant  throughout  the 
season  if  the  pump  is  operated  at  a  constant  speed  and,  consequently,  the 
gravity  head  is  generally  less  than  one-tenth  of  the  pump  head. 

Effect  of  Variations  in  Pipe  Sizes 

The  pipe  sizes  for  the  several  parts  of  the  system  selected  from  the 
tables  are  only  approximately  correct  but  the  resulting  error  should  be 
negligible  as  may  be  seen  from  the  following  study.  Assume,  as  an 
extreme  case,  that  the  error  in  pipe  size  is  so  large  that  the  water  flows 
twice  as  fast  through  one  of  the  radiators  as  through  the  others.  This 
would  make  the  friction  head  through  this  radiator  almost  four  times  as 
large  as  those  through  the  other  radiators.  The  result  would  be  that  the 
water,  in  flowing  through  the  radiator,  would  cool  5  F  instead  of  10  F. 
The  mean  water  temperature  in  the  radiator  would  then  be  217)^  F  in- 
stead of  215  F,  and  the  mean  temperature  difference,  water  to  air,  would 
be  147J^  F  instead  of  145  F.  The  heat  dissipated  by  the  radiator  would 
therefore  be  about  2  per  cent  more  than  calculated.  It  is  evident  that 
this  difference  in  heat  dissipation  is  smaller  than  the  difference  between 

567 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


TABLE  5. 


FRICTION  HEADS  (IN  MILINCHES)   OF  CENTRAL  CIRCULAR 
DIAPHRAGM  ORIFICES  IN  UNIONS 


DIAMETER 

OP 

ORIFICES 
(INCHES) 


VELOCITY  OF  WATER  IN  PIPE  IN  INCHES  PER  SECOND 


24 


36 


%-in.  Pipe 


0.25 

1300 

2900 

5000 

11,300 

20,800 

32,000 

45,000 

0.30 

650 

1450 

2500 

5700 

10,400 

16,000 

23,000 

57,000 

0.35 

330 

740 

1300 

2900 

5200 

8000 

12,000 

26,000 

47,000 

0.40 

170 

380 

660 

1500 

2600 

4000 

6800 

13,000 

24,000 

53,000 

0.45 

185 

330 

740 

1300 

2000 

2900 

6500 

12,000 

27,000 

0.50 

155 

350 

620 

970 

1400 

3200 

5700 

13,000 

0.55 

75 

170 

300 

480 

700 

1600 

2800 

6400 

1-in.  Pipe 


0.35 

900 

2000 

3500 

7800 

14,000 

22,000 

32,000 

0.40 

460 

1000 

1800 

4000 

7200 

12,000 

17,000 

37,000 

65,000 

0.45 

270 

570 

1000 

2300 

4100 

6400 

9300 

21,000 

37,000 

0.50 

160 

330 

580 

1400 

2300 

3700 

5400 

12,000 

22,000 

50,000 

0.55 

190 

330 

750 

1300 

2200 

3000 

7000 

13,000 

28,000 

0.60 

200 

440 

800 

1300 

1800 

4200 

7400 

17,000 

0.65 

120 

260 

460 

720 

1100 

2400 

4300 

10,000 

.  Pipe 


0.45 

1000 

2250 

4000 

8900 

16,000 

25,000 

36,000 

0.50 

660 

1450 

2600 

5800 

10,400 

16,400 

23,000 

53,000 

0.55 

430 

950 

1700 

3800 

6800 

10,500 

15,000 

34,000 

60,000 

0.60 

280 

630 

1100 

2500 

4400 

6900 

10,000 

22,000 

40,000 

0.65 

190 

420 

750 

1700 

3000 

4700 

6700 

15,000 

27,000 

60,000 

0.70 

285 

510 

1150 

2000 

3100 

4500 

10,000 

18,000 

40,000 

0.75 

190 

330 

750 

1300 

2100 

3000 

6700 

12,000 

26,000 

ly^-in.  Pipe 

0.55 

850 

1900 

3300 

7400 

13,000 

21,000 

30,000 

0.60 

600 

1300 

2300 

5400 

8600 

16,800 

21,000 

50,000 

0.65 

400 

850 

1500 

3600 

7200 

10,400 

14,000 

30,000 

53,000 

0.70 

260 

600 

1100 

2600 

4400 

7000 

10,000 

21,000 

39,000 

0.75 

180 

400 

760 

1800 

3000 

5000 

7000 

14,000 

28,000 

0.80 

300 

540 

1200 

2200 

3200 

5000 

10,200 

19,000 

45,000 

0.85 

200 

380 

860 

1600 

2300 

3000 

7800 

13,000 

30,000 

$-in.  Pipe 


0.70 

890 

1850 

3500 

7400 

14,000 

22,300 

33,000 

0.80 

470 

975 

1800 

3900 

7400 

11,700 

17,000 

37,000 

0.90 

255 

560 

1000 

2200 

4200 

6500 

9500 

20,500 

38,000 

1,00 

160 

340 

610 

1320 

2520 

4000 

5800 

12,500 

23,000 

49,000 

1.10 

214 

375 

850 

1600 

2500 

3700 

7900 

14,000 

30,000 

1.20 

195 

460 

950 

1360 

1910 

4200 

8100 

16,800 

1.30 

275 

525 

980 

1375 

3100 

4400 

8850 

practically  true  in  the  tests  to  determine  the  losses  of  head  in  orifices  in  $£-inM  1-in.,  and  l&-in.  pips,  con- 
ducted by  the  Texas  Engineering  Experiment  Station,  and  also  in  the  tests  to  determine  the  losses  of  head 
in  orifices  in  4-in.,  6-in.,  and  12-in.  pipe,  conducted  by  the  Engineering  Experiment  Station  of  the  University 
of  Illinois,  (Bulletin  109,  Table  6,  p.  38,  Davis  and  Jordan). 


568 


CHAPTER  33  —  HOT  WATER  HEATING  SYSTEMS  AND  PIPING 


the  calculated  heat  losses  and  the  actual  heat  losses,  and  also  smaller  than 
the  average  difference  between  the  calculated  radiator  sizes  and  the 
nearest  stock  sizes  selected. 

GRAVITY  CIRCULATION 

For  gravity  circulation,  the  one-pipe  system  shown  in  Fig.  6  and  the 
two-pipe  direct  return  system  shown  in  Fig.  7  are  probably  in  most 
common  use. 

The  one-pipe  system  has  the  disadvantage  that  the  radiator  nearest  the 


FIG.  6.    A  ONE-PIPE  GRAVITY  CIRCULATION  SYSTEM 


,   ^ 
[mtbh! 


V»  V 

14  Mbd  1 


FIG.  7.    A  TWO-PIPE  DIRECT  RETURN  GRAVITY  CIRCULATION  SYSTEM 

boiler  is  the  only  one  which  receives  water  at  approximately  the  tem- 
perature at  which  it  leaves  the  boiler.  All  other  radiators  receive  cooler 
water  and  must  be  proportionally  increased  in  size,  so  the  total  heating 
surface  in  the  system  Is  considerably  larger  than  that  in  a  corresponding 
two-pipe  system. 

The  pipe  sizes  in  gravity  circulation  systems  may  be  varied.  As  the 
pipe  sizes  are  decreased,  the  temperature  drop  through  the  radiators, 
which  produces  circulation,  is  increased  and  it  becomes  necessary  to 
increase  the  temperature  of  the  water  leaving  the  boiler  so  that  the  mean 
temperature  in  the  radiator  remains  constant.  For  example,  Fig.  8  shows 

569 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

diagrammatically  an  elementary  heating  system  which  will  function  with 
either  l}^~in.  or  1-in.  pipe.  The  radiator  is  required  to  deliver  27  Mbhr 
and  the  circuit  consists  of  30  ft  of  pipe  and  20  elbow  equivalents. 

If  IJ^-in.  pipe  is  used,  the  system  will  operate  correctly  if  the  water 
temperatures  in  the  flow  and  return  risers  are  200  F  and  180  F,  respectively. 
The  mean  water  temperature  in  the  radiators  will  then  be  190  F  and, 
if  the  radiator  is  located  in  air  having  a  temperature  of  70  F,  the  size  of 
the  radiator  must  be  sufficient  to  deliver  27  Mbh  under  these  conditions. 

If  1-in.  pipe  is  used,  the  system  will  function  correctly  with  water  tem- 
peratures in  the  flow  and  return  risers  of  210  F  and  170  F,  or  of  200  F 
and  160  F.  In  the  first  case,  the  mean  water  temperature  is  again  190  F 
and  the  same  size  radiator  may  be  used  as  with  the  IJ^-in.  pipe,  but  the 
temperature  of  the  water  leaving  the  boiler  must  be  raised  from  200  F  to 
210  F.  In  the  second  case,  the  temperature  of  the  water  leaving  the 
boiler  is  the  same  as  for  the  1^-in.  pipe,  but  the  mean  water  temperature 


FIG.  8.    AN  ELEMENTARY  SYSTEM 

in  the  radiator  is  lowered  from  190  F  to  180  F,  and  theoretically  the  size 
of  the  radiator  should  be  increased  about  12J^  per  cent  to  deliver  the 
required  27  Mbh  (See  Table  3,  Chapter  6,  1933  GUIDE). 

This  indicates  the  extent  to  which  pipe  sizes  and  radiator  sizes  may  be 
decreased  by  increasing  the  temperatures  of  the  water  in  the  boiler,  as  is 
possible  in  closed  systems  and  in  open  systems  in  which  the  open 
expansion  tank  is  located  sufficiently  high  to  secure  a  pressure  in  the 
boiler  equal  to  that  existing  in  the  boiler  of  the  closed  system. 

Example  S.  Design  a  one-pipe  gravity  circulation  system  for  the  layout  shown  in 
Fig.  6.  Assume  that  the  main  circuit  consists  of  150  ft  of  pipe,  7  elbows,  and  one  boiler. 

Solution.  Replace  the  boiler  by  3  elbow  equivalents  and  assume  that  the  size  of  the 
main  will  be  about  2  in.  According  to  Table  6,  Column  2,  a  2-in.  elbow  is  equivalent  to 
4  ft  of  pipe,  and  the  total  equivalent  length  of  the  main  will  be  about  150  plus  40,  or 
190  ft.  Assuming  that  the  center  of  the  boiler  will  be  about  4  ft  lower  than  the  horizontal 
portion  of  the  main  and  that  the  temperature  drop  in  the  system  is  to  be  35  F,  Table  6 
may  be  used  to  determine  the  size  of  the  mains.  Note  from  Column  8,  for  a  200-ft 
length,  that  a  2-in.  main  will  supply  48  Mbh  and  a  2j^-in.  main,  75,4  Mbh.  Since  the 
system  to  be  designed  is  to  supply  66  Mbh,  a  2-in.  pipe  is  too  small  and  a  2H-in.  pipe 
too  large.  The  solution  is  to  use  some  2-in.  and  some  2J4-in.  pipe-  Since  the  2H-m.  is 
nearer  the  correct  size  than  the  2-in.,  select  2-in.  pipe  for  the  first  50  or  60  ft  out  of  the 
boiler  and  2J^-in.  for  the  remaining  pipe  back  to  the  boiler. 

Tables  7  and  8  may  be  used  to  design  the  radiator  risers  and  connections.  According 
to  Table  7,  for  12  Mbh  the  flow  riser  should  be  %  in.  and  the  return  riser  1  in.,  and  the 
riser  branches  should  be  1  in.  and  1  Ji  in,,  respectively.  Note  that  according  to  Table  8, 
both  radiator  tappings  should  be  1  in.  To  simplify  the  construction,  select  1-in.  flow 
risers  with  1-in.  riser  branches  and  1-in.  radiator  tappings.  Also  select  IMrin.  return 
risers  with  lM-in.  riser  branches,  and  1  J^-in.  radiator  tappings.  Similarly,  for  18  Mbh, 
select  1-5^-in.  flow  and  return  risers  and  riser  branches,  and  lj^-in.  radiator  tappings. 

570 


CHAPTER  33 — HOT  WATER  HEATING  SYSTEMS  AND  PIPING 

TABLE  6.    CAPACITIES  OF  MAINS  IN  Mbh,  FOR  ONE-PIPE  AND  FOR  TWO-PIPE  DIRECT 

RETURN  GRAVITY  CIRCULATION  SYSTEMS  WITH  A  TOTAL  FRICTION  HEAD 

OF  0.6  IN.,  A  TEMPERATURE  DROP  OF  35  F,  WHEN  THE  MAINS 

ARE  4  FT  ABOVE  THE  CENTER  OF  THE  BOILER 


1 

2 

3 

4 

5 

6 

7 

8 

9 

10 

n 

PIPE 
SIZE 
(INCHES) 

EQUIVALENT 
LENGTH 
OF  PIPE 
(FEETa1) 

EQUIVALENT  TOTAL  LENGTH  OP  PIPE  IN  FEET  IN  LONGEST  CIRCUIT 

75 

100 

125 

ISO 

175 

200 

250 

300      1      350 

UNIT  FRICTION  HEAD,  IN  MILINCHES 

8.0 

6.0 

4.8 

4.0 

3.4 

3.0 

2.4 

2.0 

1.7 

llA 

3.0 

43.0 

57.5 

33.0 

80.0 

27.0 

85.0 

22.2 

20.2 

18.7 

2 

4.0 

83.0 

72.0 

63.0 

57.0 

51.0 

48.0 

42.0 

38.0 

35.0 

m 

4.5 

140.0 

115.0 

100.0 

90.0 

81.5 

754 

67.2 

61.0 

56.0 

3 

5.0 

234,0 

204.0 

176.5 

160.0 

143.0 

133.0 

110.0 

107.5 

100.0 

3H 

5.5 

847.0 

SOO.O 

260.0 

236.0 

214.0 

200.0 

177.0 

160.0 

146.0 

4 

6.0 

490.0 

422.0 

870.0 

334.0 

297.0 

278.0 

248.0 

223.0 

205.0 

aApproxirnate  length  of  pipe  in  feet  equivalent  to  one  elbow  in  friction  head.  This  value  varies  with 
the  velocity. 

To  develop  a  rule  for  determining  radiator  sizes,  assume  a  system 
similar  to  that  of  Fig.  6,  in  which  the  total  temperature  drop  is  to  be  35  F 
and  which  is  equipped  with  7  radiators,  all  radiators  dissipating  equal 
quantities  of  heat.  The  mean  temperature  of  the  water  in  the  radiators 
will  be  reduced  5  F  for  each  successive  radiator.  If  the  mean  tempera- 
ture of  the  water  in  the  first  radiator  is  200  F,  the  mean  temperature  of  the 


TABLE  7.    MAXIMUM  CAPACITIES  OF  RisERsa  IN  Mbh,  AND  Velocities  of  Water  in 

Pipes  in  Inches  Per  Second  FOR  ONE-PIPE  AND  FOR  TWO-PIPE  DIRECT 

RETURN   GRAVITY  CIRCULATION  SYSTEMS  WITH  A   DROP  OF 

35  F  THROUGH  EACH  RADIATOR 


PIPE  SIZE  (INCHES) 

IST  FLOOR& 

2ND  FLOOR 

3RD  A.NB  4TH  FLOORS 

or  PIPE  (FHKTC) 

Vel.  (In.perSec.)<i 

Flow 

Return 

Flow 

Return 

l/L 

1A 

•     1.0 

5 

6.2 

1^ 

% 

M 

8.0 

% 

% 

1.5 

9 

2.3 

2.3 

W.I 

14.0 

% 

1 

12 

3.2 

2.0 

12.8 

17.1 

I 

1 

2.0 

18 

2.5 

2.5 

20 

26.0 

I 

l/*i 

21 

3.0 

2.0 

25.2 

34 

1J^ 

3.0 

26 

3.0 

3.0 

43 

55 

1J^ 

1J^ 

S4 

4.0 

2.5 

i« 

m 

3.5 

43 

3.0 

3.0 

«*This  table  is  based  on  pressure  heads  of  450,  1800,  3150,  and  4500,  respectively,  for  the  first,  second, 
third,  and  fourth  floor  radiators,  and  on  friction  heads  of  200  milinches  for  the  first  floor  radiators  and  con- 
nections, and  700  milinches  for  all  other  radiators  and  their  connections. 

bThe  riser  branches,  the  piping  which  connects  the  risers  to  the  mains,  are  to  be  one  size  larger  than  the 

oApproximate  length  of  pipes  in  feet  equivalent  to  one  elbow  in  friction  head.  This  value  varies  with 
the  velocity. 

^Velocities  apply  to  the  riser  branches. 

571 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


-water  in  the  seventh  radiator  will  be  170  F,  and,  according  to  Table  3, 
Chapter  6,  of  the  1933  GUIDE,  the  heat  dissipation  of  these  two  radiators 
will  be  to  each  other  as  868  is  to  617,  or  as  140  is  to  100,  and  therefore  if 
the  last  radiator  is  to  dissipate  as  much  heat  as  the  first,  its  size  must  be 
40  per  cent  larger. 

Example  4-  Design  a  two-pipe,  direct  return,  gravity  circulation  system  for  the  lay- 
out shown  in  Fig.  7.  Assume  that  the  main  circuit  from  the  boiler  to  the  farthest  flow 
riser  and  from  the  farthest  return  riser  back  to  the  boiler  consists  of  160  ft  of  pipe, 
6  elbows,  and  1  boiler. 

Solution.  Replacing  the  boiler  by  3  elbow  equivalents  and  assuming  that  the  largest 
size  of  the  main  will  be  about  3  in.,  the  total  equivalent  length  of  the  main  will  be  160 
plus  45,  or  205  ft.  Assuming  that  the  center  of  the  boiler  will  be  about  4  ft  lower  than  the 
horizontal  portion  of  the  main,  and  that  the  temperature  drop  will  be  35  F  for  the 
system,  the  pressure  head  caused  by  the  difference  in  weight  between  the  water  in  the 

TABLE  8.    MAXIMUM  CAPACITIES  OF  RADIATOR  CONNECTIONS  IN  Mbh,  FOR  ONE-PIPE 

AND  FOR  Two- PIPE  DIRECT  RETURN  GRAVITY  CIRCULATION  SYSTEMS  WITH 

A  TEMPERATURE  DROP  OF  35  F  THROUGH  EACH  RADIATOR 


PIPE  SIZE 

IST  FLOOR 

2ND,  3RD,  AND  4TH  FLOORS 

EQUIVALENT  LENGTH 

Flow 

Return 

Off  PIPE  (FEET&) 

Mbh 

Mbk 

yz 

y* 

1.0 

4.1 

5.9 

3^ 

% 

5.2 

7.5 

3A 

% 

1.5 

7.0 

10.5 

H 

i 

9.1 

13.0 

l 

i 

2.0 

12.5 

17.8 

l 

1M 

17,5 

23.2 

IK 

1M 

3.0 

23.  3 

33.2 

aApproximate  length  of  pipe  in  feet  equivalent  to  one  elbow  in  friction  head.  This  value  varies  with 
•the  velocity. 

flow  and  return  risers  joining  the  mains  to  the  boiler  will  be  about  0.6  in.  of  water,  or 
about  one-fortieth  of  the  pressure  head  produced  by  the  circulating  pump  selected  for  the 
system  of  Fig.  3. 

Table  6  may  be  used  to  determine  the  size  of  the  main  as  follows:  Refer  to  Column  8 
and  note  that  for  Sections  AB  and  IA,  which  supply  105.6  Mbh,  a  3-in.  pipe  is  too  large 
and  a  2M-in.  pipe  is  too  small;  hence,  select  2J^  in.  for  Section  AB  and  3  in.  for  Section 
IA.  For  Sections  BC  and  HI,  which  supply  76.8  Mbh,  a  2}^-in.  pipe  is  almost  exactly 
the  correct  size  and  is  selected  for  both  sections. 

For  the  forced  circulation  system  of  Fig.  5,  the  pressure  head  produced  by  the  circu- 
lating pump  is  used  to  force  the  water  through  the  mains  and  also  through  the  risers. 
Gravity  circulation  systems  have  two  distinct  pressure  heads.  One  is  produced  by  the 
difference  in  weight  of  the  water  in  the  flow  and  return  risers  adjacent  to  the  boiler,  and 
is  the  boiler  pressure  head,  which  in  this  case  is  0.6  in.  The  other  pressure  head  is  pro- 
duced by  the  difference  in  weight  of  the  water  in  the  flow  and  return  risers  adjacent  to 
the  radiators,  and  is  the  radiator  pressure  head.  If  the  temperature  drop  through  the 
radiators  is  about  35  F,  and  if  the  story  heights  of  the  building  are  9  ft  and  the  distance 
from  the  center  of  the  first  floor  radiator  to  the  average  level  of  the  main  is  3  ft,  the 
radiator  pressure  head  of  the  first  floor  radiator  is  about  450  milinches  and  the  pressure 
heads  of  the  radiators  on  the  upper  floor  are  1350  milinches  greater  than  those  on  the 
next  lower  floors. 

Tables  6  and  7  are  based  on  the  assumption  that  the  boiler  pressure  head  must  be 
•equal  to  the  friction  head  in  the  mains,  and  that  the  several  radiator  pressure  heads  must 
be  equal  to  the  respective  radiator  and  riser  friction  heads, 

To  design  the  radiator  risers,  use  Table  7  and  begin  with  the  set  nearest  the  boiler. 
The  first  floor  risers  must  supply  28.8  Mbh.  According  to  the  table,  lM-in.  flow  and 
return  risers  will  supply  26.0  Mbh;  if  the  return  riser  is  increased  to  1 J^  in.,  the  capacity 
will  be  increased  to  34,0  Mbh,  This  is  considerably  larger  than  necessary,  and  lM~in. 
flow  and  return  risers  are  selected.  However,  it  must  be  remembered  that  the  riser 

572 


CHAPTER  33 — HOT  WATER  HEATING  SYSTEMS  AND  PIPING 

branches,  which  are  the  connections  from  the  flow  and  return  mains  to  the  flow  and 
return  risers,  are  to  be  one  size  larger  than  the  risers. 

The  second  floor  risers  must  supply  19.2  Mbh.  According  to  the  table,  the  capacity 
of  1-in.  flow  and  return  risers  is  20.0  Mbh,  and  that  size  is  selected. 

The  third  floor  risers  must  supply  9.6  Mbh.  If  a  3^-in.  flow  and  a  %-in.  return  riser 
are  used,  the  capacity  will  be  8.0  Mbh;  if  both  risers  are  %  in.,  the  capacity  will  be 
14.0  Mbh.  The  %-in-  pipe  is  selected  for  both  risers. 

To  design  the  radiator  connections,  use  Table  8  and  note  that  for  the  first  floor 
radiator  connections  the  capacity  of  a  M-in.  flow  and  1-in.  return  is  9.1  Mbh,  and  that  of 
a  1-in.  flow  and  a  1-in.  return  is  12.5  Mbh.  The  former  is  more  nearly  the  correct  size, 
but  since  it  is  difficult  to  secure  a  good  flow  through  first  floor  radiators,  the  1-in.  flow 
and  return  connection  is  selected.  For  the  two  upper  floors,  the  capacity  of  a  %-in.  flow 
and  return  connection  is  10.5  Mbh,  and  that  size  is  used. 

As  explained  in  the  design  of  the  forced  circulation  system  of  Fig.  5,, 
the  two-pipe  direct  return  system  of  Fig.  7  will  not  function  correctly 
unless  its  four  sets  of  risers  are  balanced  among  themselves.  This  neces- 
sary balancing  is  accomplished  by  adding  resistances  to  all  risers,  except 
the  one  farthest  from  the  boiler,  equal  to  the  excess  boiler  pressure  heads- 
available  for  those  risers  above  the  boiler  pressure  head  available  for  the 
farthest  riser.  For  example,  the  first  set  of  risers  is  60  ft  nearer  the  boiler 
than  the  last  set.  Since  the  flow  and  return  mains  are  designed  for  a 
friction  head  of  3  rnilinches  per  foot  (See  Table  6,  Column  8),  the  boiler 
pressure  head  available  for  the  first  set  of  risers  is  360  rnilinches  in  excess 
of  that  available  for  the  fourth  set.  The  velocity  in  the  riser  branch  is 
3  in.  per  second  (See  Table  7)  and,  therefore,  according  to  Table  5,  an 
0.65-in.  orifice  in  a  1^-in.  union  should  be  used.  This  will  provide  a 
resistance  of  about  420  rnilinches.  In  the  same  manner  it  is  found  that 
for  the  second  set  of  risers  a  resistance  of  240  rnilinches  is  required  and 
that  an  0.70-in.  orifice  in  a  1%-in.  union  will  provide  a  resistance  of  285 
rnilinches.  For  the  third  set  of  risers,  a  resistance  of  120  rnilinches  is 
required  and  an  0.60-in.  orifice  in  a  1-in.  union  will  provide  sufficient 
resistance. 

MECHANICAL  CIRCULATION 

Circulating  pumps  for  hot  water  systems  may  be  used  to  provide  the 
motive  head  for  forced  circulation  systems  as  already  described,  or  to 
improve  the  operation  of  gravity-designed  systems.  Small  specially- 
designed  centrifugal  pumps  installed  on  a  by-pass  with  the  necessary  gate 
or  check  valves  near  the  point  where  the  return  main  enters  the  heater 
maybe  employed.  Specially-designed,  electrically-driven,  propeller-type 
circulating  pumps  or  units  may  also  be  employed.  The  latter  are  usu- 
ally installed  directly  in  the  return  main  and  are  available  for  all  com- 
mercial pipe  sizes  used  for  hot  water  heating.  The  motor  switch  may  be 
under  manual  control,  automatic  control  using  thermostatic  elements, 
or  tied  in  with  the  oil  or  gas  burner  switch  which  starts  and  stops  the 
burner.  For  large  capacities  these  units  may  be  installed  in  multiple. 

For  exceptionally  large  installations  such  as  central  heating  plants,  cir- 
culating pumps  of  the  centrifugal  single  stage  type,  having  an  average 
operating  efficiency  of  70  per  cent  against  heads  up  to  125  ft,  are  some- 
times used.  It  is  generally  advisable  to  install  the  pumps  in  duplicate  to 
provide  for  contingencies  and  to  insure  continuous  operation.  In  such 
cases  each  pump  may  be  made  equal  to  two-thirds  of  the  maximum 
capacity  required. 

573 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


EXPANSION  TANKS 

When  water  at  ordinary  temperatures  is  heated  or  cooled,  its  volume  is 
increased  or  decreased.  This  variation  in  the  volume  of  the  water  in  a 
heating  system  Is  generally  provided  for  by  means  of  an  expansion  tank 
into  which  the  water  can  flow  from  the  system  during  the  heating-up 
periods  and  from  which  it  can  flow  back  into  the  system  during  the 
cooling-down  periods. 

The  expansion  tank  may  be  open  or  closed.  In  an  open  expansion  tank 
(Fig.  9),  the  water  is  subjected  to  atmospheric  pressure  and  can  expand 
freely  without  a  material  increase  in  pressure.  In  a  closed  expansion 
tank  (Fig.  10),  the  water  is  subjected  to  the  pressure  of  the  compressed  air 


VENT-**  , 

OVERFLOW  $,  VENT 


OVERFLOW 


GAUGE 
GLASS 


CIRCULATION 

P'lPE 


EXPANSION 
TANK 


Q — PRESSURE 
j  GAUGE 

^-  Gu.oae  VALVE 


EXPANSION 
TANK 


GATE  VALVEV 


CW.SUPPLY' 


{DATE  VALVE  To  DRAIW 
-GATE  VALVE 


PIPE 


FIG,  9.    AN  OPEN  EXPANSION  TANK 


-DRAIN 
FIG.  10.    A  CLOSED  EXPANSION  TANK 


within  the  tank,  and  as  the  water  expands,  the  volume  of  the  air  in  the 
tank  is  decreased  and  its  pressure  increased. 

The  open  expansion  tank  must  be  placed  at  a  sufficient  elevation  above 
the  highest  radiator  to  prevent  boiling  when  the  water  in  that  radiator  is 
at  the  highest  temperature  to  which  it  is  to  be  heated.  For  example,  if 
the  water  is  to  be  heated  to  225  F  on  extremely  cold  days,  the  absolute 
pressure  on  the  water  in  the  highest  radiator  must  be  at  least  19  Ib  per 
square  inch.  This  pressure  will  be  secured  if  the  open  expansion  tank  is 
located  15  ft  above  the  highest  radiator.  If  a  closed  expansion  tank  is 
used  and  is  located  30  ft  below  the  highest  radiator,  an  absolute  pressure 
of  about  32  Ib  per  square  inch  must  be  maintained  in  the  expansion  tank 
if  the  water  in  the  highest  radiator  is  to  be  heated  to  225  F  without  danger 
of  boiling. 

The  type  of  expansion  tank  used  in  a  heating  system,  whether  open  or 
closed,  has  no  influence  on  the  operation  of  the  system.  The  only  function 
performed  by  the  expansion  tank  is  to  provide  for  the  variation  in  the 
volume  of  the  water  in  the  system,  and  at  the  same  time  to  maintain  a 
sufficient  pressure  in  the  system  to  prevent  boiling  when  the  water  is  at 
the  highest  temperature  for  which  the  system  is  designed.  The  use  of  an 
expansion  tank  may  be  dispensed  with  when  the  heating  system  is 
allowed  to  float  on  the  water  system,  i.e.,  when  the  connection  between 

574 


CHAPTER  33 — HOT  WATER  HEATING  SYSTEMS  AND  PIPING 

the  heating  system  and  the  water  system  is  kept  open  so  that  the  water 
system  replaces  the  expansion  tank. 

The  capacity  of  the  expansion  tank  should  be  at  least  twice  the  in- 
crease in  volume  produced  when  the  water  in  the  system  is  heated  from 
its  normal  to  its  maximum  temperature.  When  25  gal  of  water  are  heated 
from  40  F  to  200  F,  the  volume  of  water  increases  to  26  gal.  A  safe  rule, 
therefore,  is  to  make  the  water  capacity  of  the  expansion  tank  equal  to 
10  per  cent  of  the  capacity  of  the  heating  system. 

In  a  forced  circulation  system,  the  expansion  tank  should  be  connected 
to  the  return  main  near  the  circulating  pump.  In  a  gravity  circulation 
system,  the  expansion  tank  should  be  connected  to  the  flow  riser  so  that 
air  liberated  from  the  water  in  the  boiler  may  escape  through  the  ex- 
pansion tank,  except  where  it  is  desired  to  maintain  a  temperature  higher 
than  212  F,  in  which  case  the  connection  should  be  in  the  return  main  to 
prevent  possible  boiling  in  the  expansion  tank. 

The  expansion  tank  should  be  protected  so  that  the  water  in  the  tank 
or  in  the  connecting  pipe  lines  cannot  freeze.  If  such  water  should  freeze 
and  the  water  in  the  system  be  heated  to  cause  further  expansion,  the 
resulting  force  will  burst  the  boiler  or  some  other  portion  of  the  system. 

INSTALLATION  DETAILS 

The  detailed  installation  of  the  pipe  system  should  be  governed  by 
four  fundamental  rules: 

1.  All  piping  must  be  pitched  either  up  or  down  so  that  all  gases  which  are  liberated 
from  the  water  can  move  freely  to  a  vented  section  of  the  system.    Whenever  practicable, 
the  pipe  line  should  be  pitched  so  that  gases  flowing  to  a  vent  will  flow  in  the  same  direc- 
tion as  the  water.    When  a  pipe  system  cannot  be  installed  without  creating  air  pockets, 
that  is,  sections  in  the  system  from  which  liberated  gases  cannot  escape,  such  sections 
must  be  provided  with  automatic  air  relief  valves  or  with  air  valves  which  may  be 
operated  manually  when  necessary. 

2.  All  piping  must  be  arranged  so  that  the  entire  system  can  be  drained,  either  to 
permit  alterations  or  repairs,  or  to  prevent  freezing  if  the  system  is  not  to  be  operated 
during  a  cold  period. 

It  is  well  to  install  a  gate  valve  and  union  in  every  riser  near  the  main  to  permit  the 
draining  of  individual  risers  without  draining  the  entire  system.  It  is  also  well,  in  large 
installations,  to  divide  the  system  into  branches  and  to  provide  each  branch  with  unions 
and  valves  so  that  any  one  branch  can  be  drained  without  disturbing  the  remaining 
ones. 

The  dividing  of  large  heating  systems  into  branches  or  zones  and  providing  each  zone 
with  individual  valves  has  the  further  advantage  of  permitting  a  varying  temperature 
control.  For  example,  if  a  building  is  equipped  with  a  forced  circulating  system  and  if 
the  south  rooms  are  on  one  branch  of  the  main  and  the  north  rooms  are  on  a  separate 
branch,  the  valves  may  be  set  so  that  the  water  will  circulate  through  the  north  branch 
with  a  temperature  drop  of,  say,  10  F,  and  through  the  south  branch  with  a  tempera- 
ture drop  of,  say,  20  F,  thus  delivering  less  heat  to  the  south  rooms  than  to  the  north 
rooms.  This  arrangement  is  especially  valuable  when  the  regulating  valves  are  controlled 
thermostatically  by  the  temperatures  in  the  two  zones,  because  no  matter  how  accurately 
the  heating  system  may  have  been  designed,  the  heat  demand  of  any  group  of  rooms 
varies  with  sunshine  and  with  wind  velocity,  and  these  intermittent  variations  can  be 
provided  for  only  by  the  individual  control  made  possible  by  changing  the  valve  settings 
controlling  the  heat  supplied  to  particular  groups  of  rooms. 

3.  All  piping  must  be  installed  so  that  it  is  free  to  expand  and  contract  with  changes  of 
temperature  without  producing  undue  stresses  in  the  pipes  or  connections.    For  this 
purpose  it  is  generally  sufficient  to  allow  for  a  variation  in  length  of  1  in.  for  100  ft  of  pipe. 

575 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

4.  The  pipe  system  must  be  installed  so  that  each  circuit  has  its  correct  friction  head. 
To  bring  this  about,  it  is  necessary  in  some  cases  to  minimize  the  friction,  i.e.,  to  make 
the  pipe  line  as  short  as  possible  and  to  provide  as  few  fittings  as  possible;  and  in  other 
cases  it  is  necessary  to  increase  the  length  of  the  pipe  and  the  number  of  fittings  so  that, 
for  every  circuit,  the  friction  head  will  be  equal  to  the  available  pressure  head. 

The  connections  from  the  boiler  to  the  mains  should  be  short  and  direct,  to  reduce  the 
friction  head.  It  is  frequently  possible  to  avoid  an  elbow  and  to  reduce  the  length  of  the 
pipe  by  running  the  pipe  in  a  diagonal  direction,  either  in  a  horizontal  or  in  a  vertical 
plane. 

The  mains  and  branches  should  pitch  up  and  away  from  the  heater,  generally  not 
less  than  1  in.  in  10  ft.  The  flow  main  should  always  be  covered;  the  return  main  should 
be  covered  except  where  it  is  to  provide  the  heating  surface  for  the  basement. 

The  connections  from  mains  to  branches  and  to  risers  should  be  such  that  circulation 
through  the  risers  will  start  in  the  right  direction.  Hence,  in  a  one-pipe  system  the  flow 
connection  must  be  nearer  the  heater  than  the  return  connection.  In  a  correctly- 
designed  two-pipe  system,  the  pressure  in  the  flow  main  is  higher  than  that  in  the  return 


FIG.  11.    METHOD  OF  CONNECTING  RADIATOR 
TO  ALLOW  FOR  EXPANSION  OF  PIPE 


main,  and  a  slight  variation  in  the  distances  of  the  flow  and  return  connections  from  the 
heater  is  not  material;  but  it  is  generally  best  to  have  the  two  connections  about  equally 
distant  from  the  heater. 

In  some  cases  it  may  be  advisable  to  take  the  flow  connection  off  the  top  of  the  main 
and  the  return  connection  from  the  side,  but  in  most  cases  both  connections  should  be  at 
an  angle  of  45  deg.  This  method  shortens  the  lines  and  substitutes  45-deg  ells  for 
90-deg  ells. 

Preferably,  connection  of  the  flow  riser  to  a  radiator  should  be  to  the  upper  tapping, 
and  connection  of  the  return  riser  to  a  radiator  should  be  to  the  lower  tapping.  When 
hot  water  enters  at  the  top  of  a  radiator  it  will  distribute  itself  along  the  entire  length  of 
the  radiator,  and  as  it  cools  it  will  settle  gradually  to  the  bottom;  the  cool  water  may 
then  be  taken  out  of  the  radiator  at  either  end. 

With  forced  circulation  and  high  velocities,  it  is  advisable  to  let  the  water  enter  at  the 
top  of  the  radiator  and  leave  at  the  bottom  of  the  opposite  end.  With  gravity  circulation 
and  low  velocities  it  makes  little  difference  whether  the  water  leaves  at  the  end  at  which 
it  enters  or  at  the  opposite  end. 

The  connections  of  the  risers  to  the  radiators  should  be  such  that  provision  is  made  for 
the  vertical  expansion  of  the  risers.  This  can  be  accomplished  as  indicated  in  Fig,  11  by 
using  one  tee  and  two  ells  for  each  connection.  These  connections  should  be  pitched 
upward  or  downward,  whichever  may  be  necessary  to  prevent  the  formation  of  air 
pockets  and  to  permit  draining. 


576 


CHAPTER  33 — HOT  WATER  HEATING  SYSTEMS  AND  PIPING 

PROBLEMS  IN  PRACTICE 

1  •  What  causes  the  circulation  of  water  in  hot  water  heating  systems? 

In  gravity  systems,  circulation  is  caused  by  the  difference  between  the  weight  of  the  cool 
water  in  the  return  riser  and  that  of  the  hot  water  in  the  flow  riser. 

In  forced  circulation  systems,  circulation  is  produced  primarily  by  a  pump,  and  second- 
arily by  the  difference  in  the  weights  of  the  water  in  the  return  and  flow  risers.  However, 
the  secondary  effect  is  so  small  when  compared  with  that  of  the  circulating  pump  that 
it  may  be  neglected  in  most  cases. 

2  •  What  tends  to  prevent  or  to  retard  the  circulation  of  water  in  hot  water 
heating  systems? 

In  both  gravity  flow  and  forced  circulation  systems,  the  friction  which  must  be  overcome 
when  the  water  is  flowing  through  pipes,  fittings,  valves,  heaters,  and  radiators  tends  to 
prevent  or  retard  circulation.  For  a  given  pipe  the  friction  varies  approximately  as  the 
1.7  power  of  the  velocity,  and  for  given  fittings,  valves,  heaters,  and  radiators,  the  friction 
varies  approximately  as  the  square  of  the  velocity.  It  is  therefore  sufficiently  accurate 
to  express  the  friction  in  fittings,  valves,  heaters,  and  radiators  in  terms  of  the  friction 
in  one  standard  elbow,  as  shown  in  Table  1. 

3  •  In  the  elementary  heating  system,  Fig.  8,  what  is  the  pressure  head  main- 
taining the  circulation  if  the  water  in  the  return  riser  is  at  180  F  and  that  in 
the  flow  riser  is  at  200  F? 

It  is  found,  from  Table  8,  Chapter  1,  that  180  F  water  weighs  60.61  Ib  per  cu  ft  and  200  F 
water  weighs  60.13  Ib  per  cu  ft.  The  pressure  head  is  independent  of  the  size  of  the 
pipe.  If  the  two  risers  were  each  1  ft  square,  the  water  in  the  flow  riser  would  weigh 
601.3  Ib  and  that  in  the  return  riser  would  weigh  606.1  Ib.  Thus  the  water  in  the  return 
riser  would  weigh  4,8  Ib  more  than  that  in  the  flow  riser.  Consequently,  the  resulting 
pressure  head  is  4.8  Ib  per  square  foot. 

Pressure  heads  are  generally  expressed  in  feet,  or  inches,  or  milinches  of  water  of  a  given 
temperature.  In  this  case  we  are  dealing  with  water  at  both  180  F  and  200  F,  so  the 
pressure  head  is  expressed  in  terms  of  190  F  water.  Such  water  weighs  60.39  Ib  per  cu  ft, 
and  to  secure  a  pressure  of  4.8  Ib  per  square  foot,  it  is  necessary  to  have  a  column  of 
water  having  a  weight  of  4.8  divided  by  60.39  =  0.0795  ft,  or  0.9540  in.,  or  954  milinches. 
This  is  the  pressure  head  which  maintains  the  circulation. 

4  •  In  the  elementary  system  of  Question  3,  if  the  radiator  dissipates  14,000 
Btu  per  hour,  what  is  the  velocity  of  the  water  in  the  pipe  line,  if  the  pipes  are 
1  in.  in  diameter?    What,  if  they  are  %  in.  in  diameter? 

Since  the  temperature  drop  through  the  radiator  is  from  200  F  to  180  F  or  20  F,  every 
pound  of  water  flowing  through  the  radiators  delivers  20  Btu;  consequently,  14,000 
divided  by  20  =  700  Ib  of  water,  or  for  190  F  water,  700  divided  by  60.39  »  11.59  cu  ft 
of  water  must  flow  through  the  radiator  and  through  the  pipe  lines  every  hour. 

The  interior  area  of  a  1-in.  pipe  is  0.864  sq  in.  The  velocity  in  the  1-in.  pipe  is  11.59 
divided  by  0.864  and  multiplied  by  144  =  1932  ft  per  hour  or  6.44  in.  per  second. 

For  %-in.  pipe,  the  interior  area  is  0.533,  and  the  velocity  is  6.44  multiplied  by  0.864 
and  divided  by  533  —  10.44  in.  per  second. 

5  •  If,  in  the  elementary  heating  system  of  Question  3,  a  1-in.  pipe  line  is 
used,  what  would  be  the  friction  head? 

If  the  radiator  is  connected  as  shown  in  Fig.  11,  with  the  heater  connected  to  provide 
freedom  of  expansion,  the  heating  circuit  may  be  assumed  to  consist  of  a  heater,  25  ft  of 
pipe,  8  elbows,  1  radiator  valve,  and  1  radiator.  From  Table  1  it  appears  that  the  heater 
and  radiator  are  equivalent,  in  friction,  to  6  elbows;  hence,  the  circuit  may  be  placed 
equal  to  25  ft  of  pipe  and  14  elbows. 

From  the  diagram  of  Fig.  4  it  appears  that  the  friction  head  for  a  1-in.  pipe  and  a  velocity 

577 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

of  6.44  in.  per  second  is  about  25  milinches  per  foot.  For  25  ft  of  pipe,  the  friction  head 
will  be  625  milinches. 

vz 
It  appears  from  Table  1  that  the  friction  head  in  one  elbow  is  —  ,  or  in  this  case  0.54 

•^g 

multiplied  by  0.54  and  divided  by  64.4  «  0.0045  ft  or  54  milinches.  Hence,  for  the  14 
elbows  the  friction  is  756  milinches.  For  the  entire  circuit,  the  friction  head  is  the  sum 
of  the  625  milinches  of  the  pipe  plus  the  756  milinches  of  the  elbows,  or  1381  milinches 
which  equal  1.381  in. 

6  •  If  the  elementary  heating  system  of  Question  3  is  installed  with  a  1-in. 
pipe  line,  how  will  it  function? 

It  is  found  from  the  answer  to  Question  3  that  the  pressure  head  is  954  milinches  and 
from  the  answer  to  Question  5  that  the  friction  head  is  1381  milinches  when  the  water  is 
flowing  with  such  velocity  that  the  specified  14,000  Btu  will  be  delivered  with  a  20  F 
temperature  drop  through  the  radiators.  Since  the  pressure  head  is  smaller  than  the 
friction  head,  the  system  will  not  function  as  planned  for  the  water  will  flow  through  the 
system  more  slowly  and  remain  in  the  radiator  longer.  The  temperature  drop  through 
the  radiator  will  be  more  than  20  F,  and  the  difference  in  the  weight  of  the  water  in  the 
return  and  flow  risers  will  be  greater  than  that  intended.  The  final  result  will  be  that  the 
pressure  head  will  become  equal  to  the  friction  head  at  a  value  somewhere  between 
954  and  1381  milinches.  Since  the  average  water  temperature  in  the  radiator  will  be 
less  than  190  F,  the  radiator  should  be  larger  than  the  size  given  in  Question  4. 

7  •  Should  a  hot  water  heating  system  be  designed  to  embody  small  pipes  or 
large  pipes? 

As  pipe  sizes  in  gravity  circulation  heating  are  reduced,  the  friction  head  is  increased- 
and  it  is  necessary  to  increase  the  temperature  drop  through  radiators;  this  lowers  the 
average  temperature  of  the  water  in  the  radiators  and  necessitates  an  increase  in  the 
size  of  the  radiators,  so  whereas  the  cost  of  the  pipe  in  a  system  is  reduced,  the  cost  of  the 
radiators  is  increased.  For  each  installation  there  is  a  definite  pipe  size  which  entails 
maximum  economy. 

As  pipe  sizes  in  forced  circulation  systems  are  reduced,  friction  heads  are  increased  so  a 
circulating  pump  of  greater  size  or  capacity  is  required.  Thus,  by  decreasing  the  size  of 
the  piping,  both  the  first  cost  of  the  circulating  pump  and  the  cost  of  its  operation  are 
increased.  There  is  a  definite  pipe  size  for  every  installation  which  is  most  economical. 
For  each  installation  of  both  types  of  systems  there  is  a  definite  pipe  size  entailing  maxi- 
mum economy  which  can  be  determined  by  a  series  of  comparative  calculations. 

8  •  What  should  be  the  size  of  the  radiators  for  the  elementary  heating  system 
of  Question  3  in  which  the  water  enters  the  radiator  with  a  temperature  of 
200  F  and  leaves  with  a  temperature  of  180  F?    The  average  temperature  of  the 
water  in  the  radiator  is,  approximately,  190  F. 

If  test  results  are  available  for  the  particular  radiators  to  be  used,  and  for  the  tempera- 
tures named,  the  size  of  the  radiators  should  be  selected  from  them.  If  no  such  test 
results  are  to  be  had,  but  if  test  results  are  available  for  the  type  of  radiator  to  be  used 
when  it  is  supplied  with  215  F  steam  and  placed  in  a  70  F  room,  the  required  size  may  be 
determined  by  the  following  ratio:  The  required  size  is  to  the  corresponding  steam 
radiator  size  as  (215  —  70)1-3  is  to  (190  -  70)1-3.  This  ratio  works  out  to  1.28.  Hence, 
the  radiators  should  be  28  per  cent  larger  under  the  conditions  prescribed  than  are  cor- 
responding radiators  under  standard  conditions.  It  is  immaterial  whether  a  radiator  is 
filled  with  steam  or  with  water,  as  long  as  the  average  temperature  of  its  outer  surface 
is  the  same  in  both  cases. 


578 


Chapter  34 

PIPE,  FITTINGS,  WELDING 

Pipe  Material,   Types  of  Pipe  Used,  Dimensions  of  Pipe  Com- 
mercially  Available,    Expansion   and   Flexibility   of  Pipe,    Pipe 
Threads  and  Hangers,  Types  of  Fittings,  Welding  as  Applied  to 
Erection  of  Piping,   Valves,  Corrosion  of  Piping 

IMPORTANT  considerations  in  the  selection  and  installation  of  pipe 
and  fittings  for  heating,  ventilating,  and  air  conditioning  work  are 
dealt  with  in  this  chapter. 

MATERIALS 

Use  of  corrosion-resistant  materials  for  pipe,  including  special  alloy 
steels  and  irons,  wrought-iron,  copper  and  brass,  has  increased  con- 
siderably during  the  past  few  years.  The  recent  development  of  copper, 
brass,  and  bronze  fittings  which  can  be  assembled  by  soldering  or  sweating 
permits  the  use  of  thin-wall  pipe  and  thereby  has  reduced  the  initial  cost 
of  such  installation.  The  following  brief  discussion  indicates  the  variety 
of  pipe  materials  and  the  types  of  pipe  available. 

Wrought-Steel  Pipe.  Because  of  its  low  price,  the  great  bulk  of  wrought 
pipe  used  for  heating  and  ventilating  work  at  the  present  time  is  of 
wrought  steel.  The  material  used  for  steel  pipe  is  a  mild  steel  made  by 
the  acid-bessemer,  the  open-hearth,  or  the  electric-furnace  process. 
Ordinary  wrought-steel  pipe  is  made  either  by  shaping  sheets  of  metal 
into  cylindrical  form  and  welding  the  edges  together,  or  by  forming  or 
drawing  from  a  solid  billet.  The  former  is  known  as  welded  pipe,  the 
latter  as  seamless  pipe. 

Many  types  of  welded  pipe  are  available,  although  the  smaller  sizes 
most  frequently  used  in  heating  and  ventilating  work  are  made  by  the 
lap-weld  or  butt-weld  process.  While  the  lap-weld  process  produces  a 
better  weld  than  the  butt  type,  lap-weld  pipe  is  seldom  manufactured  in 
nominal  pipe  sizes  less  than  2  in.  Seamless  pipe  can  be  obtained  in  the 
small  sizes  at  a  somewhat  higher  cost. 

Seamless  steel  pipe  is  frequently  used  for  high  pressure  work  or  where 
pipe  is  desired  for  close  coiling,  cold  bending,  or  other  forming  operation. 
Its  advantages  are  its  somewhat  greater  strength  which  permits  use  of  a 
thinner  wall  and,  in  the  small  sizes,  its  freedom  from  the  occasional 
tendency  of  welded  pipe  to  split  at  the  weld  when  bent. 

Wrought-iron  Pipe.  Wrought-iron  pipe  is  considered  to  be  more  corro- 
sion-resisting than  ordinary  steel  pipe  and  therefore  its  somewhat  higher 
first  cost  can  be  justified  on  the  basis  of  longer  life  expectancy.  Wrought- 
iron  pipe  may  be  identified  by  the  spiral  line  marked  into  each  length, 
either  knurled  into  the  metal  or  painted  on  it  in  red  or  other  bright  color. 

579 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

Otherwise,  there  is  little  difference  in  the  appearance  of  wrought-iron  and 
steel  pipe,  although  microscopic  examination  of  polished  and  etched 
specimens  will  readily  disclose  the  difference. 

Cast  Ferrous  Pipe.  There  are  now  available  several  types  of  cast 
ferrous-metal  pipe  made  of  a  good  grade  of  cast-iron  with  or  without 
additions  of  nickel,  chromium,  or  other  alloy.  This  pipe  is  available  in 
sizes  from  1 J^  in.  to  6  in.,  and  in  standard  lengths  of  5  or  6  ft  with  external 
and  internal  diameters  closely  approximating  those  of  extra-strong 
wrought  pipe.  Cast  ferrous  pipe  may  be  obtained  coupled,  bevelled  for 
welding,  or  with  ends  plain  or  grooved  for  the  several  types  of  couplings. 
It  is  easily  cut  and  threaded  as  well  as  welded.  The  fact  that  it  is  readily 
welded  enables  the  manufacturers  to  supply  the  pipe  in  any  lengths 
practicable  for  handling. 

Alloy  Metal  Pipe.  Steel  pipe  bearing  a  small  alloy  of  copper  or  other 
alloying  element,  such  as  molybdenum  or  manganese,  has  been  claimed 
to  possess  more  resistance  to  corrosion  than  plain  steel  pipe  and  it  is 
advertised  and  sold  under  various' trade  names. 

Copper  Pipe  and  Fittings.  Owing  to  its  inherent  resistance  to  cor- 
rosion, copper  and  brass  pipe  have  always  been  used  in  heating,  venti- 
lating, and  water  supply  installations,  but  the  cost  with  standard  dimen- 
sions for  threaded  connections  has  been  high.  The  recent  introduction 
of  fittings  which  permit  erection  by  soldering  or  sweating  allows  the  use 
of  pipe  with  thinner  walls  than  are  possible  with  threaded  connections, 
thereby  reducing  the  cost  of  installations. 

The  initial  cost  of  brass  and  copper  pipe  installations  generally  runs 
higher  than  the  corresponding  job  with  steel  pipe  and  screwed  connections 
in  spite  of  the  use  of  thin-wall  pipe,  but  the  corrosive  nature  of  the  fluid 
conveyed  or  the  inaccessibility  of  some  of  the  piping  may  warrant  use  of 
a  more  expensive  material  than  plain  steel.  The  advantages  of  corrosion- 
resisting  pipe  and  fittings  should  be  weighed  against  the  correspondingly 
higher  initial  cost. 

DIMENSIONS 

The  /PS  dimensions  of  commercial  pipe  universally  used  at  the  present 
time  conform  to  the  recommendations  made  by  a  Committee  of  the 
A.S.M.E.  in  1886.  Pipe  up  to  12  in,  in  diameter  is  made  in  certain 
definite  sizes  designated  by  nominal  internal  diameter  which  is  somewhat 
different  from  the  actual  internal  diameter,  depending  on  the  wall  thick- 
ness required.  There  are  three  weights  of  wrought-iron  and  steel  pipe 
commonly  used,  known  as  standard-weight,  extra-strong,  and  double-extra- 
strong.  Because  of  the  necessity  of  maintaining  the  same  external  dia- 
meter in  all  three  weights  for  the  same  nominal  size,  the  added  wall 
thickness  is  obtained  by  decreasing  the  internal  diameter.  The  term 
full-weight,  when  applied  to  sizes  below  8  in,,  means  that  the  pipe  is  up  to 
the  nominal  weight  per  foot.  When  applied  to  sizes  between  8  and  12  in., 
inclusive,  it  often  indicates  that  the  pipe  has  the  heaviest  of  several  wall 
thicknesses  listed.  In  sizes  14  in.  and  upward,  pipe  is  designated  by  its 
outside  diameter  (O.D.)  and  the  wall  thickness  is  specified. 

While  the  demands  for  pipe  for  the  heating  and  ventilating  industry  are 
reasonably  well  served  by  the  standard-weight  and  extra-strong  pipe, 

580 


CHAPTER  34 — PIPE,  FITTINGS,  WELDING 


demands  for  pipe  for  higher  pressures  and  temperatures  in  industry 
resulted  in  the  use  of  a  multiplicity  of  wall  thicknesses  for  all  sizes.  Even 
in  heating  installations,  the  erection  of  piping  by  welding  was  deemed  to 
warrant  the  use  of  pipe  lighter  than  standard  weight.  For  these  reasons, 
a  Sectional  Committee  on  Standardization  of  Wrought  Iron  and  Wrought 
Steel  Pipe  and  Tubing  functioning  under  the  procedure  of  the  American 
Standards  Association  was  appointed  to  standardize  the  dimensions  and 
materials  of  pipe. 

The  proposed  pipe  standard  recommended  by  that  sectional  committee 
has  set  up  several  schedules  of  pipe  including  standard-weight  and  extra- 
strong  thicknesses  which  are  now  included  in  Schedules  40  and  60,  re- 
spectively. The  schedules  approved  by  the  Sectional  Committee  are 
given  in  Tables  1  and  3  and  the  corresponding  weights  in  Tables  2  and  4. 

Standard-weight  pipe  is  generally  furnished  with  threaded  ends  in 
random  lengths  of  16  to  22  ft,  although  when  ordered  with  plain  ends, 

5  per  cent  may  be  in  lengths  of  12  to  16  ft.    Five  per  cent  of  the  total 
number  of  lengths  ordered  may  be  jointers  which  are  two  pieces  coupled 
together.     Extra-strong  pipe  is  generally  furnished  with  plain  ends  in 
random  lengths  of  12  to  22  ft,  although  5  per  cent  may  be  in  lengths  of 

6  to  12  ft. 

EXPANSION  AND  FLEXIBILITY 

The  increase  in  temperature  of  a  pipe  from  room  temperature  to  an 
operating  steam  or  water  temperature  one  hundred  degrees  or  more  above 
room  temperature  results  in  an  increase  in  length  of  the  pipe  for  which 
provision  must  be  made.  The  amount  of  linear  expansion  (or  contraction 
in  the  case  of  refrigeration  lines)  per  unit  length  of  material  per  degree 
change  in  temperature  is  termed  the  coefficient  of  linear  expansion  of 
that  material,  or  commonly,  the  coefficient  of  expansion.  This  coefficient 
varies  with  the  material. 

The  linear  expansion  of  cast  iron,  steel,  wrought-iron,  and  copper  pipe, 
the  materials  most  frequently  used  in  heating  and  ventilating  work,  can 
be  determined  from  Table  5. 

The  elongation  values  in  Table  5  were  computed  from  the  following 
formula : 


1000 ")  +  H"1000"  '    I  (1) 


•where 


Lt  =  length  at  temperature  t  degrees  Fahrenheit,  feet. 
Z0  »  length  at  32  F,  feet. 

t  =  final  temperature,  degrees  Fahrenheit. 
a  and  b  are  constants  as  follows: 


METAL 

a 

6 

Cast-  Iron 

0.005441 

0.001747 

Steel     .  ,  .,. 

0.006212 

0.001623 

Wrought-  Iron              .         ... 

0.006503 

0.001622 

CoDoer  „    

0.009278 

0.001244 

581 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


The  three  methods  by  which  the  elongation  due  to  thermal  expansion 
may  be  taken  care  of  are: 

1.  Expansion  joints. 

2.  Swivel  joints. 

3.  Inherent  flexibility  of  the  pipe  itself  utilized  through  pipe  bends,  right-angle  turns, 
or  offsets  in  the  line. 

Expansion  joints  of  the  slip-sleeve,  diaphragm,  or  corrugated  types 
made  of  copper,  rubber,  or  other  gasket  material  are  all  used  for  taking 

TABLE  1.    DIMENSIONS  OF  WELDED  AND  SEAMLESS  STEEL  PIPE 


NOMINAL 
PIPE  SIZE 

OUTSIDE 

DJAM. 

NOMINAL  WA.LL  THICKNESSES  FOR  SCHEDULE  NUMBERS 

Schedule 
10 

Schedule 
20 

Schedule 
30 

Schedule 

4:0 

Schedule 
60 

Schedule 
80 

Schedule 
100 

Schedule 
120 

Schedule 
140 

Schedule 
160 

y* 

H 

H 

ilA 
Vi 

2 

m 

3 

m 

4 
5 
6 
8 
10 
12 
14  O.  D. 
160.D. 
18  O.  D. 
20  0.  D. 
24  0.  D. 
30  O.  D. 

0.405 
0.540 
0.675 
0.840 
1.050 
1.315 
1.660 
1,900 
2.375 
2.875 
3.500 
4.000 
4.500 
5.563 
6.625 
8.625 
10.75 
12.75 
14.0 
16.0 
18.0 
20.0 
24.0 
30.0 

0.068* 

0.095* 

0.088* 

0.119* 

0.091* 

0.126* 

0.109* 

0.147* 

0.187 
0.218 

0.250 
0,250 
0.281 
0.343 
0.375 
0.437 

0.113* 

0.154* 

0.133* 

0.179* 

0.140* 

0.191* 

0.145* 

0.200* 

0.154* 

0.218* 

0.203* 

0.276* 

0.216* 

0.300* 

0.226* 

0.318* 

0.237* 

0.337* 

0.437 

0.531 
0.625 
0.718 
0.906 
1.125 
1.312 
1.406 
1.562 
1.750 
1.937 
2.312 

0.258* 

0.375* 

0.500 

0.280* 

0.432* 

0.562 

O."25(f 
0.250 
0.250 
0.250 
0.250 
0.312 

0.250 
0.250 
0.250 
0.312 
0,312 
0.312 
0.375 
0.375 
0.500 

0.277* 
0.307* 
0.330* 
0.375 
0.375 
0.437 
0.500 
0.562 
0.625 

0.322* 
0.365* 
0.406 
0.437 
0.500 
0.562 
0.593 
0.687 

0.406 
0.500* 
0.562 
0.593 
0.656 
0.718 
0.812 
0.937 

0.500* 
0.593 
0.687 
0.750 
0.843 
0.937 
1.031 
1.218 

0.593 

0.718 
0.843 
0.937 
1.031 
1.156 
1.250 
1.500 

0.718 
0.843 
1.000 
1.062 
1.218 
1.343 
1.500 
1.750 

0.812 

1.000 
1.125 
1.250 
1.437 
1.562 
1.750 
2.062 

All  dimensions  are  given  in  inches, 

The  decimal  thicknesses  Hated  for  the  respective  pipe  sizes  represent  their  nominal  or  average  wall  dimensions  and  include 
an  allowance  for  mill  tolerance  of  12.5  per  cent  under  nominal  thicknesses. 

Thicknesses  marked  with  asterisk  in  Schedules  30  and  40  are  identical  with  thicknesses  for  standard-weight  pipe  in 
former  lists;  those  in  Schedules  60  and  80  are  identical  with  thicknesses  for  extra-strong  pipe  in  former  lists. 

The  Schedule  Numbers  indicate  approximate  values  of  the  expression  1000  x  P/S. 

up  expansion,  but  generally  only  for  low  pressures  or  where  the  inherent 
flexibility  of  the  pipe  cannot  readily  be  used  as  in  underground  steam  or 
hot  water  distribution  lines. 

Swivel  joints  are  used  extensively  in  low-pressure  steam  and  hot  water 
heating  systems  and  in  hot  water  supply  lines.  The  swivel  joints  absorb 
the  expansive  movement  of  the  pipe  by  the  turning  of  threaded  joints. 
In  many  cases  the  straight  pipe  in  the  offset  of  a  swivel  joint  is  sufficiently 
flexible  to  take  up  the  expansion  without  developing  enough  thrust  to 
produce  swiveling  in  the  threaded  joint.  This  is  preferable  since  con- 
tinued turning  in  the  threaded  joint  may  in  time  result  in  a  leak,  par- 

582 


CHAPTER  34 — PIPE,  FITTINGS,  WELDING 


ticularly  when  the  pressure  is  high.  The  amount  of  elongation  which  a 
swivel  joint  can  take  up  is  controlled  by  the  length  of  the  swing  piece 
employed  and  by  the  lateral  displacement  which  is  permissible  in  the 
long  pipe  runs. 

Probably  the  most  economical  method  of  providing  for  expansion  of 
piping  in  a  long  run  is  to  take  advantage  of  the  directional  changes  which 
must  necessarily  occur  in  the  piping  and  proportion  the  offsets  so  that 
sufficient  flexibility  is  secured.  Ninety-degree  bends  with  long,  straight 
tangents  in  either  a  horizontal  or  a  vertical  plane  are  an  excellent  means 

TABLE  2.    NOMINAL  WEIGHTS  OF  WELDED  AND  SEAMLESS  STEEL  PIPE 


NOMINAL 

PIPE 

SlZB 

INCHES 

SCHED. 

10 
PLAIN 

ENDS 

SCHED. 
20 
PLAIN 

ENDS 

SCHEDULE 
30 

SCHEDULE 
40 

SCHED. 
60 
PLAIN 
ENDS 

SCHED. 
80 
PLAIN 

ENDS 

SCHED. 
100 

PLAIN 

ENDS 

SCHED. 
120 
PLAIN 

ENDS 

SCHED. 
140 
PLAIN 
ENDS 

SCHED. 
160 
PLAIN 

ENDS 

Plain 
Ends 

Threads 
and 
Coup- 
lings 

Plain 
Ends 

Threads 
and 
Coup- 
lings 

y* 
¥ 

H 
l 

M 

VA 
i 

m 

3 

m 

4 
5 
6 
8 
10 
12 
14  0.  D. 
160.D. 
18  0.  D. 
20  0.  D. 
240.  D. 
30  0.  D. 

0.25* 
0.43* 
0.57* 
0.86* 
1.14* 
1.68* 
2.28* 
2.72* 
3.66* 
5.80* 
7.58* 
9.11* 
10.8* 
14.7* 
19.0* 
28.6* 
40.5* 
53.6 
63.3 

0.25* 
0.43* 
0.57* 
0.86* 
1.14* 
1.69* 
2.29* 
2.74* 
3.68* 
5.82* 
7.62* 
9.21* 
10.9* 
14.9* 
19.2* 
28.8* 
41.2* 
55.0 

0.32* 

0.54* 

0.74* 

r~LTL_ 

1.09* 

1.31 
1.94 
2.85 
3.77 
4.86 
7.45 
10.0 
14.3 

1.48* 



2.18* 

3.00* 

3.64* 

5.03* 

7.67* 

10.3* 

12.5* 

15.0* 

19.0 

22.6 
33.0 
45.3 
74.7 
116.0 
161.0 
190.0 
241.0 
304.0 
374.0 
536.0 

20.8* 

27.1 

28.6* 

36.4 

22.4 
28.1 
33.4 
45.7 
52.3 
59.0 
78.6 
94.7 
158.0 

24.7* 
34.3* 
43.8* 
54.6 
62.6 
82.0 
105.0 
141.0 
197.0 

25.0* 
35.0* 
45.0* 

35.7 

54.8* 
73.2 
85.0 
108.0 
133.0 
167.0 
231.0 

43.4* 
64.4 
88.6 
107.0 
137.0 
171.0 
209.0 
297.0 

50.9 
77.0 
108.0 
131.0 
165.0 
208.0 
251.0 
361.0 

60.7 
89.2 
126.0 
147.0 
193.0 
239.0 
297.0 
416.0 

67.8 
105.0 
140.0 
171.0 
224.0 
275.0 
342.0 
484.0 

lisT 

42.1 
47.4 
52.8 
63.5 
99.0 

82.8 

105.0 

123.0 

171.0 

Weights  are  given  in  pounds  per  linear  foot  and  are  for  pipe  with  plain  ends  except  for  sizes  which  are  commercially  available  with 
threads  and  couplings  for  which  both  weights  are  listed. 

"The  weights  marked  with  asterisk  in  Schedules  30  and  40  are  identical  with  weights  for  standard~weight  pipe  in  former  lists:  those  in 
Schedules  60  and  80  are  identical  with  weights  for  extras-strong  pipe  in  former  lists. 

The  Schedule  Numbers  indicate  approximate  values  of  the  expression  1000  x  P/S. 

for  securing  adequate  flexibility  with  larger  sizes  of  pipe.  When  flexi- 
bility cannot  be  obtained  in  this  manner,  it  is  necessary  to  make  use  of 
some  type  of  expansion  bend.  The  exact  calculation  of  the  size  of  ex- 
pansion bends  required  to  take  up  a  given  amount  of  thermal  expansion 
is  relatively  complicated1,  The  following  approximate  method,  however, 
has  been  found  to  give  reasonably  good  results  and  is  deemed  to  be 
sufficiently  accurate  for  most  heating  work. 
Fig.  3,  Chapter  32,  shows  several  types  of  expansion  bends  commonly 

ipiping  Handbook,  by  Walker  and  Croker,  and  A  Manual  for  the  Design  of  Piping  for  Flexibility  by 
the  Use  of  Graphs,  by  E,  A.  Wert,  S.  Smith,  and  E.  T.  Cope,  published  by  The  Detroit  Edison  Company, 

583 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


used  for  taking  up  thermal  expansion.    The  amount  of  pipe,  L,  required 
in  each  of  these  bends  may  be  computed  from  the  following  formula: 

(2) 


L  =  6.16  JDA 


where 

L  =  length  of  pipe,  feet. 

D  =  outside  diameter  of  the  pipe  used,  inches. 

A-  =f  the  amount  of  expansion  to  be  taken  up,  inches. 

This  formula,  based  on  the  use  of  mild-steel  pipe  with  wall  thicknesses 
not  heavier  than  extra-strong,  assumes  a  maximum  safe  value  of  fiber 
stress  of  16,000  Ib  per  square  inch.  When  square  type  bends  are  used,  the 
width  of  the  bend  should  not  exceed  about  two  times  the  height.  It  is 

*-  TABLE  3.    DIMENSIONS  OF  WELDED  WROUGHT-IRON  PIPE 


NOMINAL 
PIPE 
SIZE 

OUTSIDE 
DIAMETER 

NOMINAL  WALL  THICKNESSES  FOR  SCHEDULE  NUMBERS 

Schedule 
10 

Schedule 
20 

Schedule 
30 

Schedule 
40 

Schedule 
60 

Schedule 
80 

1A 
1A 

y* 
IA 

H 

m 
VA 

2 

m 

3 
"  3H 
4 
5 
6 
8 
10 
12 
140.  D. 
16  0.  D. 
18  0.  D. 
20  0.  D. 

0.405 
0.540 
0.675 
0.840 
1.050 
1.315 
1.660 
1.900 
2.375 
2.875 
3.5 
4.0 
4.5 
5.563 
6.625 
8.625 
10.75 
12.75 
14.0 
16.0 
18.0 
20.0 

0.070* 
0.090* 
0.093* 
0.111* 
0.115* 
0.136* 
0.143* 
0.148* 
0.158* 
0.208* 
0.221* 
0.231* 
0.242* 
0.263* 
0.286* 
0.329* 
0.372* 
0.414 
0.437 
0.500 
0.562 
0.562 

0.098* 
0.122* 
0.129* 
0.151* 
0.157* 
0.183* 
0.195* 
0.204* 
0.223* 
0.282* 
0.306* 
0.325* 
0.344* 
0.383* 
0.441* 
0.510* 
0.606 
0.702 
0.750 

0.283* 
0.313* 
0.336* 
0,375 
0.375 
0,437 
0.500 

0.510* 
0.574 
0.625 
0.6S7 
0.750 

0.250 
0,250 
0.250 

0.312 
0.312 
0.312 
0.375 

All  dimensions  are  given  in  inches. 

The  decimal  thicknesses  listed  for  the  respective  pipe  sizes  represent  their  nominal  or  average  wall  dimensions  and  include 
an  allowance  for  mill  tolerance  of  12.5  per  cent  under  the  nominal  thickness. 

•"Thicknesses  marked  with  an  asterisk  in  Schedules  30  and  40  are  identical  with  thicknesses  for  standariL-weight  pipe  in 
former  lists;  those  in  Schedules  60  and  80  are  identical  with  thicknesses  for  extra-strong  pipe  in  former  lists. 

The  Schedule  Numbers  indicate  approximate  values  of  the  expression  1000  x  P/S. 

further  assumed  that  the  corners  are  made  with  screwed  or  flanged  elbows 
or  with  arcs  of  circles  having  radii  five  to  six  times  the  pipe  diameter. 

All  risers  must  be  anchored  and  safeguarded  so  that  the  difference  in 
length  when  hot  from  the  length  when  cold  shall  not  disarrange  the 
normal  and  orderly  provisions  for  drainage  of  the  branches, 

It  is  especially  necessary  with  light-weight  radiators  so  to  anchor  the 
piping  and  so  to  give  it  freedom  for  expansion  that  no  strain  therefrom 
shall  be  allowed  to  distort  the  radiators.  When  expansion  strains  from 
the  pipes  are  permitted  to  reach  these  light  metal  heaters  they  usually 
emit  sounds  of  distress  which  are  exceedingly  troublesome. 

584 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


supports  fitted  with  rollers,  machined  blocks,  elliptical  or  circular  rings  of 
larger  diameter  than  the  pipe  giving  contact  only  at  the  bottom  or  trolley 
hangers.  In  all  cases,  allowance  should  be  made  for  rod  clearance  to 
permit  swinging  without  setting  up  severe  bending  action  in  the  rods. 
For  pipes  of  small  size,  perforated  metal  strip  is  often  used.  For 
horizontal  mains,  the  rod  or  strip  usually  is  attached  to  the  joists  or  steel 

TABLE  5.   THERMAL  EXPANSION  OF  PIPE  IN  INCHES  PER  100  Fxa 
(For  superheated  steam  and  other  fluids  refer  to  temperature  column) 


SATURATED  STEAM 

ELONGATION  IN  INCHES  PER 
100  FT  PROM  —  20  F  UP 

SATURATED 
STEAM 

ELONGATION  IN  INCHES  PER  100 
FT  FROM  —20  F  UP 

Vacuum 
Inches 
ofHg. 

Pressure 
Pounds 
per 
Square 
Inch 
Gage 

Tem- 
perature 
Degrees 
Fahren- 
heit 

Cast- 
iron 
Pipe 

Steel 
Pipe 

Wrought 
Iron 
Pipe 

Copper 
Pipe 

Pressure 
Pounds 
per 
Square 
Inch 
Gage 

Tem- 
perature 
Degrees 
Fahren- 
heit 

Cast- 
iron 
Pipe 

Steel 
Pipe 

Wrought 
Iron 
Pipe 

Copper 
Pipe 

-20 

0 

0 

0 

0 

664.3 

500 

3.847 

4.296 

4.477 

6.110 

0 

0.127 

0.145 

0.152 

0.204 

795.3 

520 

4.020 

4.487 

4.677 

6.352 

20 

0.255 

0.293 

0.306 

0.442 

945.3 

540 

4.190 

4.670 

4.866 

6.614 

40 

0.390 

0.430 

0.465 

0.655 

1115.3 

560 

4.365 

4.860 

5.057 

6.850 

29.39 

60 

0.518 

0.593 

0.620 

0.888 

1308.3 

580 

4.541 

5.051 

5.268 

7.123 

28.89 

80 

0.649 

0.725 

0.780 

1.100 

1525.3 

600 

4.725 

5.247 

5.455 

7.388 

27.99 

100 

0.787 

0.898 

0.939 

1.338 

1768.3 

620 

4.896 

5.437 

5.660 

7.636 

26.48 

120 

0.926 

1.055 

1,110 

1.570 

2041.3 

640 

5.082 

5.627 

5.850 

7.893 

24.04 

140 

1.051 

1.209 

1.265 

1.794 

2346.3 

660 

5.260 

5.831 

6.067 

8.153 

20.27 

160 

1.200 

1.368 

1.427 

2.008 

2705 

680 

5.442 

6.020 

6.260 

8.400 

14.63 

180 

1.345 

1.528 

1.597 

2.255 

3080 

700 

5.629 

6.229 

6.481 

8.676 

6.45 

200 

1.495 

1.691 

1.778 

2.500 

720 

5.808 

6.425 

6.673 

8.912 

"Ts" 

220 

1.634 

1.852 

1.936 

2.720 

740 

6.006 

6.635 

6.899 

9.203 

10.3 

240 

1.780 

2.020 

2.110 

2.960 

760 

6.200 

6.833 

7.100 

9.460 

20.7 

260 

1.931 

2.183 

2.279 

3.189 

780 

6,389 

7.046 

7.314 

9.736 

34.5 

280 

2.085 

2.350 

2.465 

3.422 

800 

6.587 

7.250 

7.508 

9.992 

52.3 

300 

2.233 

2.519 

2.630 

3.665 

820 

6.779 

7.464 

7.757 

10.272 

74.9 

320 

2.395 

2.690 

2.800 

3.900 

840 

6.970 

7.662 

7.952 

10.512 

103.3 

340 

2.543 

2.862 

2.988 

4.145 

860 

7.176 

7.888 

8.195 

10.814 

138.3 

360 

2.700 

3.029 

3.175 

4.380 

880 

7.375 

8.098 

8.400 

11.175 

180.9 

380 

2.859 

3.211 

3.350 

4.628 

900 

7.579 

8.313 

8.639 

11.360 

232.4 

400 

3.008 

3.375 

3.521 

4.870 

920 

7.795 

8.545 

8.867 

11.625 

293.7 

420 

3.182 

3.566 

3.720 

5.118 

940 

7.989 

8.755 

9.089 

11.911 

366.1 

440 

3.345 

3.740 

3.900 

5.358 

960 

8.200 

8.975 

9.300 

12.180 

451.3 

460 

3.511 

3.929 

4.096 

5.612 

980 

8.406 

9.196 

9,547 

12.473 

550.3 

480 

3.683 

4.100 

4.280 

5.855 

1000 

8.617 

9.421 

9.776 

12.747 

aFrom  Piping  Handbook,  by  Walker  and  Crocker.  This  table  gives  the  expansion  from  —20  F  to  the 
temperature  in  question.  To  obtain  the  amount  of  expansion  between  any  two  temperatures  take  the 
difference  between  the  figures  in  the  table  for  those  temperatures,  For  example,  if  a  steel  pipe  is  installed 
at  a  temperature  of  60  F  and  is  to  operate  at  300  F,  the  expansion  would  be  2.519  -  0.593  «•  1.926  in. 

work  of  the  floor  above.  For  long  runs  of  vertical  pipe  subject  to  con- 
siderable thermal  expansion,  either  the  hangers  should  be  designed  to 
prevent  excessive  load  on  the  bottom  support  when  expansion  takes 
place,  or  the  bottom  support  should  be  designed  to  withstand  the  entire 
load. 

FITTINGS 

Fittings  for  joining  the  separate  lengths  of  pipe  together  are  made  in  a 
variety  of  forms,  and  are  either  screwed  or  flanged,  the  former  being 

586 


•CHAPTER  34 — PIPE,  FITTINGS,  WELDING 


generally  used  for  the  smaller  sizes  of  pipe  up  to  and  including  3J^  in., 
and  the  latter  for  the  larger  sizes,  4  in.  and  above.  Screwed  fittings  of 
large  size  as  well  as  flanged  fittings  of  small  size  are  also  made  and  are 
used  for  certain  classes  of  work  at  the  proper  pressure. 

The  material  used  for  fittings  is  generally  cast-iron,  but  in  addition  to 
this  malleable  iron,  steel  and  steel  alloys  are  also  used,  as  well  as  various 

TABLE  6.    TENTATIVE  AMERICAN  STANDARD  DIMENSIONS  OF  ELBOWS,  45  DEC  ELBOWS, 
TEES,  AND  CROSSES  (STRAIGHT  SIZES)  FOR  125  LB  CAST-IRON  SCREWED  FITTINGS 


ELBOW 


TEE 


CROSS 


A 

c 

B 

E 

F 

G 

H 

CENTER 

INSIDE  DIAMETER 

NOMINAL 
PIPE 
SIZE 

TO  END, 
ELBOWS, 
TEES  AND 
CROSSES 

CENTER 
TO  END, 
45   DEG 
ELBOWS 

LENGTH 
OP  THREAD 

MIN. 

WIDTH 
OF  BAND, 

MIN. 

OF  FITTING 

METAL 
THICKNESS. 
MIN. 

OUTSEDB 

DIA.MBTBH 
OF  BAND, 

MIN. 

Min. 

Max. 

X 

0.81 

0.73 

0.32 

0.38 

0.540 

0.584 

0.110 

0.93 

0.95 

0.80 

0.36 

0.44 

0.675 

0.719 

0.120 

1.12 

% 

1.12 

0.88 

0.43 

0.50 

0.840 

0.897 

0.130 

1.34 

% 

1.31 

0.98 

0.50 

0.56 

1.050 

1.107 

0.155 

1.63 

1 

1.50 

1.12 

0.58 

0.62 

1.315 

1.385 

0.170 

1.95 

1M 

1.75 

1.29 

0.67 

0.69 

1.660 

1.730 

0.185 

2.39 

1H 

1.94 

1.43 

0.70 

0.75 

1.900 

1.970 

0.200 

2.68 

2 

2.25 

1.68 

0.75 

0.84 

2.375 

2.445 

0.220 

3.28 

2H 

2.70 

1.95 

0.92 

0.94 

2.875 

2.975 

0.240 

3.86 

3 

3.08 

2.17 

0.98 

1.00 

3.500 

3.600 

0.260 

4.62 

3H 

3.42 

2.39 

1.03 

1.06 

4.000 

4.100 

0.280 

5.20 

4 

3.79 

2.61 

1.08 

1.12 

4.500 

4.600 

0.310 

5.79 

5 

4.50 

3.05 

1.18 

1.18 

5.563 

5.663 

0.380 

7.05 

6 

5.13 

3.46 

1.28 

1.28 

6.625 

6-725 

0.430 

8.28 

8 

6.56 

4.28 

1.47 

1.47 

8.625 

8.725 

0.550 

10.63 

10 

8.08 

5.16 

1.68 

1.68 

10.750 

10.850 

0.690 

13.12 

12 

9.50 

5.97 

1.88 

1.88 

12.750 

12.850 

0.800 

15.47 

14  O.D. 

10.40 

.... 

2.00 

2.00 

14.000 

14.100 

0.880 

16.94 

16  O.D. 

11,82 

.... 

2.20 

2.20 

16.000 

16.100 

1.000 

19.30 

All  dimensions  given  in  inches. 

grades  of  brass  or  bronze.     The  material  to  be  used  depends  on  the 
character  of  the  service  and  the  pressure. 

As  in  the  case  of  pipe,  there  are  several  weights  of  fittings  manufactured. 
Recognized  American  Standards  for  the  various  weights  are  as  follow: 

Cast-iron  pipe  flanges  and  flanged  fittings  for  25  Ib  (sizes  4  in.  and  larger),  125  Ib,  and 
250  Ib  maximum  saturated  steam  pressure. 

587 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


Malleable  iron  screwed  fittings  for  150  Ib  maximum  saturated  steam  pressure. 
Cast-iron  screwed  fittings  for  125  and  250  Ib  maximum  saturated  steam  pressure. 
Steel  flanged  fittings  for  150  and  300  Ib  maximum  steam  service  pressure. 
The  allowable  cold  water  working  pressures  for  these  standards  vary  from  43  Ib  for 
the  25  Ib  standard  to  500  Ib  for  the  300  Ib  steel  standard. 

Screwed  fittings  include :  nipples  or  short  pieces  of  pipe  of  varying 
lengths;  couplings,  usually  of  wrought  iron  only ;  elbows  for  turning  angles 
of  either  45  deg  or  90  deg;  return  bends,  which  may  be  of  either  the  close 
or  open  pattern,  and  may  be  cast  with  either  a  back  or  side  outlet;  tees; 
crosses;  laterals  or  Y  branches;  and  a  variety  of  plugs,  bushings,  caps, 

TABLE  7.    AMERICAN  STANDARD  DIMENSIONS  OF  TEES  AND  CROSSES  (STRAIGHT  SIZES) 
FOR  125  LB  CAST-IRON  FLANGED  FITTINGS 


n 


K-A-*t  U-A-»t*-A-*l 

^fen    fefe! 


A 

IJL 


TEE 


SIDE  OUTLET 


CROSS 


NOMINAL 
PIPE  SlZEa-b 

A 

AA 

DIAMETER 

OF 

FLANGE 

THICKNESS  OF 
FLANGE, 
MIN. 

METAL 
THICKNESS 
OF  BODY, 
MIN. 

CENTER  TO  FACE 
TEES  AND 
CROSSES  *>•« 

FACE  TO  FACE 
TEES  AND 
CROSSES  b-c 

1 

IK 

3% 

7 

|g 

ge 

He 

4 

8  2 

5  8 

%.  6 

$ 

2  2 

4H 

9 

6 

/"S 

2J-£ 

5 

10 

7 

l^Hlg 

7/a 

3 

5Ji 

11 

7J^ 

% 

P* 

3-^5 

6 

12 

81/g 

ljj/£g 

4 

6Jx£ 

13 

9 

Ijjj^g 

% 

5 

7J/£ 

15 

10 

*JK6 

l^ 

6 

8 

16 

11 

1 

9it 

8 

9 

18 

13J^ 

1M 

10 

11 

22 

16 

Ij^g 

% 

12 

12 

24 

19 

IJi 

1^f« 

14  O.D. 

14 

28 

21 

1  % 

H 

16  O.D. 

15 

30 

23  J^ 

1/^L6 

1 

18  O.D. 

16J^ 

33 

25 

l?f  6 

JK« 

20  O.D. 

18 

36 

27  J^ 

11J^6 

24  O.D. 

22 

44 

32 

l!% 

1J^ 

30  O.D. 

25 

50 

38% 

2H 

IJie 

36  O.D, 

28 

56 

46 

2J^ 

15^ 

42  O.D. 

31 

62 

53 

25^ 

l1Me 

48  O.D. 

34 

68 

59H 

2M 

2 

All  dimensions  given  in  inches. 

aSize  of  all  fittings  listed  indicates  nominal  inside  diameter  of  port. 

bTees,  side  outlet  tees,  and  crosses,  16  in,  and  smaller,  reducing  on  the  outlet,  have  the  same  dimensions 
center  to  face,  and  face  to  face  as  straight  size  fittings  corresponding  to  the  size  of  the  larger  opening. 
Sizes  18  in,  and  larger,  reducing  on  the  outlet,  are  made  in.  two  lengths,  depending  on  the  size  of  the  outlet. 

cTees  and  crosses,  reducing  on  run  only,  carry  same  dimensions  center  to  face  and  face  to  face  as  a 
straight  size  fitting  of  the  larger  opening. 

588 


CHAPTER  34 — PIPE,  FITTINGS,  WELDING 


lock-nuts,  flanges  and  reducing  fittings.  Reducing  fittings  a?  well  as 
bushings,  both  of  which  are  used  in  changing  from  one  pipe  size  to  another, 
may  have  the  smaller  connection  tapped  eccentrically  to  permit  free  drain- 
age of  the  water  of  condensation  in  steam  lines  or  free  escape  of  air  in 
water  lines. 

Threads  used  for  fittings  are  the  same  American  Standard  taper  pipe 
threads  as  those  used  for  pipe,  and  unless  otherwise  ordered,  right-hand 
threads  are  used.  To  facilitate  drainage,  some  elbows  have  the  thread 

TABLE  8.     AMERICAN  STANDARD  DIMENSIONS  OF  ELBOWS  FOR 
125  LB  CAST-IRON  FLANGED  FITTINGS 


VU/ 

1        *        •      1 


SO  OEG.       LOHGfcADlUS      45OEG,  REDUCING      SIDE  OUTLET 


NOMINAL 
PIPE  SIZE  a 


CENTER  TO  FACE 
ELBOW  b-«-<j 


CENTER  TO  FACE 
LONG  RADIOS 
ELBOW  b-o-d 


CENTER  TO  FACE 
45    DBG 
ELBOW  « 


DIAMETER 

OF 

FLANGE 


THICKNESS 

OP  FLANGE, 

Mm. 


METAL 
THICKNESS 
OF  BODY, 

MlN. 


1 

2  2 

3  2 

4 

5 

6 

8 
10 
12 

14  O.D. 
16  O.D. 
18  O.D. 
20  O.D. 
24  O.D. 
30  O.D. 
36  O.D. 
42  O.D. 
48  0,D. 


3% 
4 


7J 

8 

9 
11 
12 
14 
15 


18 
22 
25 
28 
31 
34 


5 

i* 


9 

lOJi 
11  Ji 
14 


19 


24 

26^ 
29 
34 

41H 
49 


64 


I* 


5 
5H 

8  2 


11 
15 
18 
21 
24 


9 
10 
11 


16 

19 

21 

23^ 

25 


32 


46 
53 


All  dimensions  given  in  Inches. 

aSize  of  all  fittings  listed  indicates  nominal  inside  diameter  of  port, 

^Reducing  elbows  and  side  outlet  elbows  carry  same  dimensions  center  to  face  as  straight  size  elbows 
corresponding  to  the  size  of  the  larger  opening. 

oSpecial  degree  elbows,  ranging  from  1  to  45  deg,  inclusive,  have  the  same  center  to  face  dimensions 
as  given  for  45  deg  elbows  and  those  over  45  deg  and  up  to  90  deg,  inclusive,  shall  have  the  same  center  to 
face  dimensions  as  given  for  90  deg  elbows.  The  angle  designation  of  an  elbow  is  its  deflection  from  straight 
line  flow  and  is  the  angle  between  the  flange  faces> 

outlet  elbows  shall  have  all  openings  on  intersection  center-lines. 


580 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

tapped  at  an  angle  to  provide  a  pitch  of  the  connecting  pipe  of  J^  in.  to 
the  foot.  These  elbows  are  known  to  the  trade  as  pitched  elbows  and  are 
commercially  available.  Malleable  iron  fittings,  like  brass  fittings,  are 
cast  with  a  round  instead  of  a  flat  band  or  bead,  or  with  no  bead  at  all. 
Fittings  are  designated  as  male  or  female,  depending  on  whether  the 
threads  are  on  the  outside  or  inside,  respectively. 

Flanged  fittings  are  generally  used  in  the  best  practice  for  connecting 
all  piping  above  4  in.  in  diameter.  While  screwed  fittings  may  be  used 
for  the  larger  sizes  and  are  satisfactory  under  the  proper  working  con- 
ditions, it  will  be  found  difficult  either  to  make  or  to  break  the  joints  in 
these  large  sizes. 

A  number  of  different  flange  facings  in  common  use  are  plain  face, 
raised  face,  tongue  and  groove,  and  male  and  female.  Cast-iron  fittings 
for  125  Ib  pressure  and  below  are  normally  furnished  with  a  plain  face, 
while  the  250  Ib  cast-iron  fittings  are  supplied  with  a  J^-inch  raised  face. 
The  standard  facing  for  steel  flanged  fittings  for  150  and  300  Ib  is  a 
}^-inch  raised  face  although  these  fittings  are  obtainable  with  a  variety  of 
facings.  The  gasket  surface  of  the  raised  face  may  be  finished  smooth 
or  may  be  machined  with  concentric  or  spiral  grooves  often  referred  to  as 
serrated  face  or  phonograph  finish,  respectively. 

The  dimensions  of  elbows,  tees  and  crosses  for  125  Ib  cast-iron  screwed 
fittings  are  given  in  Table  6,  whereas  the  dimensions  for  125  Ib  cast-iron 
flanged  fittings  are  given  in  Tables  7  and  8. 

For  low  temperature  service  not  to  exceed  about  220  F,  a  number  of 
paper  or  vegetable  fiber  gasket  materials  will  prove  satisfactory ;  for  plain 
raised  face  flanges,  rubber  or  rubber  inserted,  gaskets  are  commonly 
employed.  Asbestos  composition  gaskets  are  probably  the  most  widely 
used,  particularly  where  the  temperature  exceeds  250  F.  Jacketed 
asbestos  and  metallic  gaskets  may  be  used  for  any  pressure  and  tem- 
perature conditions,  but  preferably  only  with  a  relatively  narrow  recessed 
facing. 

WELDING 

Erection  of  piping  in  heating  and  ventilating  installations  by  means  of 
fusion  welding  has  been  commonly  accepted  in  the  past  few  years  as  a 
competitive  method  to  the  screwed  and  flanged  joint.  Since  the  question 
of  economy  of  welding  as  against  the  use  of  screwed  and  flanged  fittings 
is  dependent  on  the  individual  job,  the  use  of  welding  is  generally  recom- 
mended on  the  basis  of  a  greatly  reduced  cost  of  maintenance  and  repair, 
of  less  weight  resulting  from  the  use  of  a  lighter-weight  pipe,  and  of 
increased  economy  in  pipe  insulation,  hangers,  and  supports  rather  than 
on  the  basis  of  any  economy  that  might  be  effected  in  actual  erection  by 
welding. 

Fusion  welding,  commonly  used  in  erection  of  piping,  is  defined  as  the 
process  of  joining  metal  parts  in  the  molten,  or  molten  and  vapor  states, 
without  the  application  of  mechanical  pressure  or  blows.  Fusion  welding 
embraces  gas  welding  and  electric  arc  welding,  both  of  which  are  com- 
monly used  to  produce  acceptable  welds. 

Welding  application  requires  the  same  basic  knowledge  of  design  as  do 
the  other  types  of  assembly,  but  in  addition,  requires  a  generous  know- 

590 


CHAPTER  34 — PIPE,  FITTINGS,  WELDING 


ledge  of  the  sciences  involved,  particularly  as  to  welding  qualities  of 
metal,  their  reaction  to  extremely  high  temperatures,  and  the  ability  to 
determine  and  use  only  the  best  quality  welding  rods.  This  requirement 
applies  equally  to  employer  and  employee  with  the  employer  accepting 
all  of  the  responsibility.  Thus  the  employer  should  select  his  welding 
mechanics  with  good  judgment,  provide  them  with  first-class  equipment 
and  tools,  arrange  for  their  training  and  use  of  acceptable  workmanship 
standards,  and  at  regular  intervals  subject  their  work  to  prescribed  tests. 


a.  TYPICAL  SHORT  RADIUS  ELBOWS 


b.  TEE  c.  FORGED  CAP 

FIG.  1.    TYPICAL  WELDING  FITTINGS 


d.  CONCENTRIC 
REDUCER 


e.  END 
CLOSURE 


Industry  will  not  accept  the  employment  of  mechanics  of  undetermined 
ability  nor  on  the  basis  of  past  experience.  Neither  does  industry  accept 
the  statement  that  a  weld  is  only  as  good  as  the  workman  who  makes  it. 
The  control  Codes  now  in  process  of  adoption  will  be  the  law  governing 
the  use  of  the  welding  process.  These  Codes  prohibit  individual  practices 
contrary  to  their  specified  procedure  and  rules  of  control,  and  this  is 
predicated  upon  the  sound  requirement  that  the  employer  must  assume 
full  responsibility  for  the  deposited  weld. 

It  is  advisable  that  this  management  responsibility  be  included  in  all 
welding  specifications  and  that  authoritative  standards  of  workmanship 
also  be  specified.  The  standards  of  workmanship  for  this  industry  are  as 
set  forth  in  the  Standard  Manual  on  Pipe  Welding  of  the  Heating,  Piping 
and  Air  Conditioning  Contractors  National  Association. 

A  complete  line  of  manufactured  steel  welding  fittings  is  now  available 
with  plain  ends  machine  beveled  for  welding  and  with  radii  similar  to 
short  and  long  radius  flanged  fittings.  Some  typical  types  of  these  fittings 

591 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

TABLE  9.    PROPOSED  DIMENSIONS  OF  STEEL  WELDING  NECK  FLANGES  FOR 

MAXIMUM  STEAM  SERVICE  PRESSURE  OF  150  LB  PER  SQ  IN. 

(GAGE)  AT  A  TEMPERATURE  OF  500  F,  AND  100  LB  AT  750  F 


NOMINAL 
PIPE 


DIAMETER 

OF 

FLANGE 


THICKNESS 

OF 

FLG.  MIN. 


DIAMETER 

OP 

HUB 


HUB  DIAM. 
BEGINNING 
OP  CHAMPEK 


R 


LENGTH 
THRU 
HUB 


DIAM.  FOR 

STANDARD 
PIPE 


DIAM.  OF 
BOLT 
CIRCLE 


No. 

OF 

BOLTS 


SIZE 

OP 

BOLTS 


2 
3  2 

4 

5 

6 

8 
10 
12 

14  0.  D. 
16  0.  D. 
18  O.  D. 
20  0.  D. 
24  0,  D. 


9 

10 
11 


16 
19 
21 


M 


1H 


2^6 


32! 


5%  6 
7^6 


12 
143* 


25 


32 


l^8 
frjfe 


1.32 

1.66 

1.90 

2.38 

2.88 

3.50 

4.00 

4.50 

5.56 

6.63 

8.63 

10.75 

12.75 

14.00 

16.00 

18,00 

20.00 

24.00 


05 
38 
61 
07 


2.47 


07 
55 
03 
05 
07 
98 


10.02 
12.00 
13.25 
15.25 
17.25 
19.25 
23.25 


17 

ISM 


22% 

25 


4 

4 

4 

4 

4 

4 

8 

8 

8 

8 

8 

12 

12 

12 

16 

16 

20 

20 


H 

72 


All  dimensions  given  in  inches. 

A  raised  face  of  ^  in.  is  included  in  thickness  of  flange  minimum. 

It  is  recommended  that  the  taper  of  the  hub  should  not  exceed  6  degrees  for  a  reasonable  distance  back  of  the  chamfer 
in  order  to  reduce  the  heat  transfer  while  welding. 


are  shown  in  Fig.  1.  They  are  made  in  pipe  sizes  ^  to  24  in.,  standard 
and  extra  heavy,  in  steel,  wrought  iron,  brass,  copper,  and  special  alloys. 
Socket  welding  fittings  shown  in  Fig.  1  are  commercially  available.  A 
proposed  American  Standard  containing  dimensions  of  steel  welding-neck 
flanges  for  pressures  up  to  1500  Ib  per  square  inch  has  been  developed  in 
A.S.A.  Sectional  Committee  B16.  Tables  9  and  10  give  these  dimensions 
for  welding-neck  flanges  suitable  for  150  and  300  Ib  per  square  inch 
gage  pressure. 

VALVES 

Valves  are  made  with  both  threaded  and  flanged  ends  for  screwed  and 
bolted  connections  just  as  are  pipe  fittings. 

The  material  used  for  valves  of  small  size  is  generally  brass  or  bronze 
for  low  pressures  and  forged  steel  for  high  pressures,  while  in  the  larger 

592 


CHAPTER  34 — PIPE,  FITTINGS,  WELDING 


sizes  either  cast-iron,  cast-steel  or  some  of  the  steel  alloys  are  employed. 
Practically  all  iron  or  steel  valves  intended  for  steam  or  water  work  are 
bronze-mounted  or  trimmed. 

Brass,  bronze,  and  iron  valves  are  generally  designed  for  standard  or 
extra  heavy  service,  the  former  being  used  up  to  125  Ib  and  the  latter  up 
to  250  Ib  saturated  steam  working  pressure,  although  most  manufacturers 
also  make  valves  for  medium  pressure  up  to  175  Ib  steam  working  pres- 
sure. The  more  common  types  are  gate  valves  or  straightway  valves, 
globe  valves,  angle  valves,  check  valves  and  automatic  valves,  such  as 
reducing  and  back-pressure  valves. 

Gate  valves  are  the  most  frequently  used  of  all  valves  since  in  their  open 
position  the  resistance  to  flow  is  a  minimum.  These  valves  may  be 
secured  with  either  a  rising  or  a  non-rising  stem,  although  in  the  smaller 
sizes  the  rising  stem  is  more  commonly  used.  The  rising  stem  valve  is 
desirable  because  the  positions  of  the  handle  and  stem  indicate  whether 
the  valve  is  open  or  closed,  although  space  limitations  may  prevent  its 
use.  The  globe  valve  is  less  expensive  to  manufacture  than  the  gate 
valve,  but  its  peculiar  construction  offers  a  high  resistance  to  flow  and 
may  prevent  complete  drainage  of  the  pipe  line.  These  objections  are  of 
particular  importance  in  heating  work. 

Check  valves  are  automatic  in  operation  and  permit  flow  in  only  one 
direction,  depending  for  operation  on  the  difference  in  pressure  between 
the  two  sides  of  the  valve.  The  two  principal  kinds  of  check  valves  are 
the  swing  check  in  which  a  flapper  is  hinged  to  swing  back  and  forth,  and 
the  lift  check  in  which  a  dead  weight  disc  moves  vertically  from  its  seat. 

Valves  commonly  used  for  controlling  steam  or  water  supply  to  radi- 
ators constitute  a  special  class  since  they  are  manufactured  to  meet 
heating  system  requirements.  These  valves  are  generally  of  the  angle 
type  and  are  usually  made  of  brass.  Graduations  on  the  heads  or  lever 
handles  are  often  supplied  to  indicate  the  relative  opening  of  the  valve  in 
any  position.  Standard  roughing-in  dimensions  for  angle-type  valves 
are  given  in  Table  11. 

Automatic  control  of  steam  supply  to  individual  radiators  can  be 
effected  by  use  of  direct-acting  radiator  valves  having  a  thermostatic 
element  at  the  valve,  or  near  to  it.  The  direct-acting  valve  is  usually  an 
angle-type  valve  containing  a  thermostatic  element  which  permits  the 
flow  of  steam  in  accordance  with  room  temperature  requirements.  These 
valves  usually  are  capable  of  adjustment  to  permit  variation  in  room 
temperature  to  suit  individual  taste. 

Ordinary  steam  valves  may  be  used  for  hot  water  service  by^drilling  a 
^-in»  hole  through  the  web  forming  the  seat  to  insure  sufficient  circulation 
to  prevent  freezing  when  the  valve  is  closed.  Valves  made  particularly 
for  use  in  hot  water  heating  systems  are  of  less  complex  design,  one  type 
consisting  of  a  simple  butterfly  valve,  and  another  of  a  quick  opening  type 
in  which  a  part  in  the  valve  mechanism  matches  up  with  an  opening 
in  the  valve  body. 

In  one-pipe  steam-heating  systems,  automatic  air  valves  are  required 
at  the  radiators.  Two  common  types  of  air  valves  available  are  the 
vacuum  type  and  the  straight-pressure  type.  Vacuum  valves  permit  the 

593 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


TABLE  10.    PROPOSED  DIMENSIONS  OF  STEEL  WELDING  NECK  FLANGES  FOR 

MAXIMUM  STEAM  SERVICE  PRESSURE  OF  300  LB  PER  SQ  IN. 

(GAGE)  AT  A  TEMPERATURE  OF  750  F 


NOMINAL 
PIPE 
SIZE 


DUM. 

OF 

FLANGE 


THICK- 
NESS 
OP 
FLANGE 

MIN. 


DlAM. 
OF 

HUB 


HUB 

DlAM. 

BEGINNING 

OP 
CHAMFER 


H 


LENGTH 
THRU 
HUB 


DlAM. 

FOR 

STANDARD 
PIPE 


DIAM. 

FOR 

EXTRA 
STRONG 
PIPE 


DlAM. 
OF 

BOLT 
CIRCLE 


No. 

OF 

BOLTS 


SIZE 

OF 

BOLTS 


*2 


4 

5 

6 

8 
10 
12 

14  0.  D. 
16  0.  D, 
18  0.  D. 
20  0.  D. 
24  0.  D. 


Ji 


m 

9 
10 
11 


3«e 
3% 
&A 

$H 

7 


15 


23 


28 

30H 

36 


2M 


19 
21 
23^ 


2.38 

2.88 

3.50 

4.00 

4.50 

5.56 

6.63 

8.63 

10.75 

12.75 

14.00 

16.00 

18.00 

20.00 

24.00 


2.07 

2.47 

3.07 

3.55 

4.03 

5.05 

6.07 

7.98 

10.02 

12.00 

13.25 

15.25 

17.25 

19.25 

23.25 


1.94 


32 
90 
36 

83 
81 
76 
7.63 
9.75 
11.75 


13 


22^ 


27 
32 


8 

8 

8 

8 

8 

8 

12 

12 

16 

16 

20 

20 

24 

24 

24 


*For  sizes  below  2  inches  use  dimensions  of  600  Ib  flanges. 
All  dimensions  given  in  inches. 

A  raised  face  of  fa  in.  is  included  in  thickness  of  flange  minimum. 

It  is  recommended  that  the  taper  of  the  hub  should  not  exceed  6  degrees  for  a  reasonable  distance  back  of  the  chamfer 
in  order  to  reduce  the  heat  transfer  while  welding. 


expulsion  of  air  from  the  radiators  when  the  steam  pressure  rises  and,  in 
addition,  act  as  checks  to  prevent  the  return  of  air  into  the  radiator  when 
a  vacuum  is  formed  by  the  condensation  ofsteam  after  the  supply  pressure 
has  dropped.  Ordinary  air  valves  permit  the  expulsion  of  air  from  the 
radiator  when  steam  is  supplied  under  pressure,  but  when  the  pressure 
dies  down  and  a  vacuum  tends  to  be  formed  the  air  is  drawn  back  into 
the  radiator. 

A  system  supplied  with  vacuum  valves  will  heat  more  quickly  and 
stay  warm  longer  than  one  provided  with  straight  pressure  air  valves; 
thus  it  will  effect  considerable  economy  of  fuel  because  the  idle  period 
during  which  no  heat  is  delivered  is  shortened.  Automatic  air  valves  are 
provided  with  a  float  to  close  them  in  case  the  radiator  becomes  flooded 
with  water  because  it  does  not  drain  properly. 

594 


CHAPTER  34 — PIPE,  FITTINGS,  WELDING 


CORROSION2 

Corrosion  is  sometimes  encountered  in  heating  work  on  the  outside  of 
buried  pipes  or  the  inside  of  steam  heating  systems;  it  is  seldom  ex- 
perienced in  hot  water  heating  systems  unless  the  water  is  frequently 
renewed.  Piping  buried  in  the  ground  is  quite  successfully  protected  by 
coatings  of  the  asphaltic  type  which  are  usually  applied  hot  and  often 
reinforced  with  fabric  wrappings.  Galvanizing  by  the  hot-dip  process  and 
painting  with  specially  prepared  mixtures  also  afford  some  protection. 

Internal  corrosion  in  steam  heating  systems  occurs  principally  in  the 
condensate  return  pipes  and  is  nearly  always  caused  by  oxygen  or  carbon 
dioxide,  or  both,  in  solution  in  the  condensate.  Oxygen  may  enter  the 
heating  system  with  the  steam,  owing  to  its  presence  in  the  boiler-feed 
water,  or  it  may  enter  as  air  through  small  leaks,  particularly  in  systems 
which  operate  at  sub-atmospheric  pressures.  When  a  steam  heating 
system  is  operated  intermittently,  air  rushes  in  during  each  shutdown 
period  and  oxygen  is  absorbed  by  the  condensate  which  clings  to  the 
interior  surfaces  of  the  pipes  and  radiators.  The  rate  of  corrosion  depends 
upon  the  amounts  of  oxygen  and  carbon  dioxide  present  in  solution,  upon 
the  operating  temperature,  and  upon  the  length  of  time  that  the  pipe 
surfaces  are  in  contact  with  gas-laden  condensate. 

Another  possible  cause  of  corrosion  is  a  flow  of  electric  current  some- 
times resulting  from  faulty  electrical  circuits  which  should  be  corrected. 
Electrolytic  corrosion  also  may  occur  because  of  the  presence  of  two  dis- 
similar metals,  such  as  brass  and  iron,  but  the  condensate  in  practically 
all  steam  heating  systems  is  such  a  weak  electrolyte  that  this  cause  of 
corrosion  is  very  infrequent. 

If  trouble  is  experienced  from  corrosion,  oxygen  should  be  eliminated 
from  the  feed  water  by  proper  deaeration  with  commercial  apparatus. 
The  elimination  of  the  oxygen  due  to  air  leakage  is  more  difficult  because 
of  the  multitude  of  small  leaks  which  exist  around  valve  stems  and  in 
pipe  joints.  In  vacuum  systems,  however,  an  attempt  should  be  made 
to  minimize  such  leakage. 

Carbon  dioxide  in  varying  amounts  is  contained  in  steam  produced 
from  the  majority  of  water  supplies.  It  is  formed  from  the  breaking  down 
of  carbonates  and  bicarbonates  which  are  present  in  nearly  all  natural 
waters.  It  can  be  partly  removed  by  chemical  treatment  and  deaeration, 
but  there  is  no  simple  method  whereby  it  can  be  entirely  eliminated. 

These  gases  cause  corrosion  only  when  in  solution  in  the  condensate; 
when  they  are  mixed  with  dry  steam  their  corrosive  effect  is  negligible. 
The  amount  of  gas  in  solution  depends  upon  the  partial  pressure  of  that 
gas  in  the  atmosphere  above  the  surface  of  the  solution,  in  accordance 
with  the  well  known  physical  law  of  Henry  and  Dalton8.  The  exact 
application  of  this  law,  however,  assumes  equilibrium  conditions  which 
do  not  always  exist  under  the  flow  conditions  prevailing  in  a  heating 
system. 


*New  Light  on  Heating  System  Corrosion,  by  J.  H.  Walker  (Heating  and  Ventilating,  May,  1933). 
"Some  Fundamental  Considerations  of  Corrosion  in  Steam  and  Condenaate  Lines,  by  R.  E.  Hall  and 
A.  R.  Mumford  (A.S.H.V.E.  TRANSACTIONS,  Vol.  38,  1932). 

595 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

Distinction  should  be  made  between  corrosion  in  heating  systems  proper 
and  in  the  condensate  discharge  lines  from  other  apparatus  using  steam, 
such  as  water  heaters,  kitchen  equipment,  and  sterilizers.  Experience 
has  shown  that  in  heating  systems  the  partial  pressures  of  the  gases  do 
not  reach  such  magnitudes  as  to  cause  harmful  amounts  of  gas  to  become 
dissolved  in  the  condensate  when  steam  supplies  are  of  reasonable  purity. 
In  other  kinds  of  steam-using  apparatus  which  are  not  ordinarily  well 
vented,  the  gases  tend  to  accumulate  in  the  steam  space  and  to  become 
dissolved  in  the  condensate' in  appreciable  concentrations.  Consequently, 
corrosion  is  frequently  observed  in  the  condensate  discharege  lines  from 
such  apparatus,  but  this  does  not  necessarily  indicate  that  equally  serious 
corrosion  is  taking  place  in  the  heating  system  supplied  with  steam  from 
the  same  source. 

TABLE  11.    STANDARD  ROUGHING-IN  DIMENSIONS  ANGLE  TYPE  VALVES 


SIZE 

OF 

VALVH 

DIMENSION  A 
STEAM  AND 
Hor  WATER  ANGLE  VALVES 
AND  UNION  ELBOWS 
EFFECTIVE  JANUARY  1,  1926 

DIMENSION  A 
MODULATING  VALVES 
EFFECTIVE  JANUARY  1,  1926 

DIMENSION  A 
RETURN  LINE  VACUUM 
VALVES  EFFECTIVE 
JANUARY  1,  1925 

1  4 

3  4 

2$t 
3 

?* 

2 
Tolerance 

3H 

3j| 

All  dimensions  given  in  inches. 

Connecting  ends  shall  be  threaded  and  gaged  as  to  threading  according  to  the  American  (Taper)  Pipe  Thread  Standard, 
A.S.A.  No.  B2—1919. 

The  standardization  of  the  Roughing-in  Dimensions  of  Angle  Steam  and  Hot  Water,  and  Modulating  Radiator  Valves  was 
made  possible  by  the  cooperation  of  the  Manufacturers  Stanaardiaation  Society  of  the  Valves  and  Fittings  Industry. 


When  corrosive  conditions  are  believed  to  exist,  their  seriousness  should 
be  determined  by  actual  measurement,  rather  than  by  inference  from 
isolated  instances  of  pipe  failures.  The  National  District  Heating  Associa- 
tion has  perfected  a  corrosion  tester  for  measuring  the  inherent  corrosive- 
ness  of  existing  conditions.  This  corrosion  tester  consists  of  a  frame  sup- 
porting three  coils  of  wire  which  are  carefully  weighed.  After  the  tester 
has  been  inserted  in  the  pipe  line  for  a  definite  length  of  time,  the  loss  of 

596 


CHAPTER  34 — PIPE,  FITTINGS,  WELDING 


weight  of  the  coils,  referred  to  an  established  scale,  indicates  the  relative 
corrosiveness  of  the  condensate.  Accompanying  such  corrosion  measure- 
ments, a  careful  chemical  analysis  should  be  made  of  the  condensate,  and 
the  findings  will  serve  as  a  basis  for  an  intelligent  study  of  the  problem. 

Corrosion,  if  found  to  exist,  can  be  lessened  or  overcome  by  several 
means.  If  the  steam  supply  is  found  to  be  definitely  contaminated,, 
proper  chemical  treatment  of  the  water,  followed  by  deaeration,  is  an 
obvious  remedy.  The  leaks  in  the  piping  system,  particularly  in  vacuum 
systems,  should  be  stopped  so  far  as  is  practicable. 

Some  success  has  been  reported  with  the  use  of  inhibitors,  chief  among 
which  are  oil,  sodium  silicate,  and  ammonia.  Oil  may  be  fed  into  the 
main  steam-supply  pipe  by  means  of  a  sight-feed  lubricator.  The  type  of 
oil  known  as  600- W  is  usually  recommended.  In  the  present  state  of 
knowledge  on  this  point,  the  quantity  to  be  fed  can  best  be  determined  by 
trial.  The  use  of  sodium  silicate,  fed  in  a  similar  manner,  is  reported  to 
be  successful  but  it  has  not  been  widely  used. 

The  effect  of  ammonia  is  to  increase  the  pH  value  of  the  condensate 
above  the  point  where  corrosion  is  likely  to  take  place.  Speller4  reports 
having  injected  small  quantities  of  ammonia  into  a  small  closed  heating 
system  (the  entire  amount  of  condensate  being  returned  to  the  boiler)  and 
finding  the  pH  value  maintained  at  a  high  point  for  several  months 
without  further  additions  of  ammonia.  The  concentration  of  ammonia 
must  be  kept  low  to  avoid  corrosion  of  brass  parts  of  the  system.  The  use 
of  ammonia  is  not  to  be  recommended  where  steam  may  come  in  contact 
with  food  or  other  materials. 

In  view  of  the  fact  that  corrosion  is  most  frequently  found  in  the 
return  lines  from  special  equipment,  which  constitute  a  relatively  small 
part  of  the  total  piping  in  a  building,  a  simple  solution  of  the  corrosion 
problem  may  be  to  use  non-corroding  materials  in  those  certain  portions 
of  the  piping  system,  since  the  higher  cost  will  usually  be  an  unappreciable 
portion  of  the  total.  Brass  and  copper  are  undoubtedly  less  subject  to 
this  type  of  corrosion  than  the  ferrous  metals,  and  considerable  attention 
is  now  being  given  to  corrosion-resistant  linings  for  ferrous  pipe.  Cast- 
iron  pipe,  sometimes  alloyed  with  other  metals,  also  deserves  con- 
sideration. 


^Corrosion  in  Steam  Heating  Systems,  by  F,  N.  Speller  (A.S.H.V.E.  TRANSACTIONS,  Vol.  34,  1928). 


PROBLEMS  IN  PRACTICE 

1  •  What  materials  are  used  for  pipes  of  heating  systems? 

Steel  pipe  is  generally  used  for  steam  piping  and  for  return  lines  where  corrosion  is  not 
particularly  active.  Where  corrosion  is  an  important  factor,  it  is  good  practice  to  use 
special  corrosion-resistant  ferrous  alloys,  or  brass  or  copper  pipe. 

2  •  Why  is  thin-walled  copper  pipe  made  up  with  sweated  joints? 

If  the  pipe  were  threaded  it  would  be  necessary  to  use  at  least  standard-weight  wall 
thickness  on  account  of  the  metal  removed  in  threading.  Flared  ends  with  coupling  nuts 
may  be  used,  but  this  construction  is  expensive  and  hard  to  keep  tight. 

597 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

3  •  How  are  pipes  designated  in  diameters  of  12  in.  and  less? 

By  weight  and  nominal  size,  referring  to  the  approximate  inside  diameter. 

4  •  How  are  pipe  sizes  designated  in  diameters  of  14  in.  and  more? 

By  wall  thickness  and  outside  diameter. 

5  •  Why  are  expansion  joints  required  in  steam  pipes? 

To  care  for  the  change  in  length  of  the  line  brought  about  by  a  change  in  temperature, 

6  •  What  devices  are  used  for  taking  up  expansion? 

Expansion  joints,  swivel  joints,  and  the  inherent  flexibility  of  the  pipe  itself. 

7  •  Where  are  swivel  joints  principally  used? 

In  branch  connections  to  radiators,  and  in  the  risers  of  multi-story  buildings  where  they 
are  installed  between  the  floor  joists. 

8  •  Name  three  grades  of  American  Standard  screwed  pipe  fittings. 

125-lb  cast-iron,  150-lb  malleable  iron,  and  250-lb  cast-iron. 

9  •  In  wfyat  sizes  are  American  standard  cast-iron  flanges  and  flanged  fittings 
for  25-lb  saturated  steam  pressure  made? 

In  nominal  sizes  from  4  in.  to  72  in.,  inclusive. 


598 


Chapter  35 

WATER  SUPPLY  PIPING 

Maximum    Possible    Flow,    Maximum    Probable    Flow,    Average 

Probable  Flow,   Factor  of   Usage,   Kind  of  Pipe  Used,   Sizing  of 

Risers,  Sizing  of  Mains,  Sizing  of  Systems,  Hot   Water  Supply, 

Hot  Water  Storage 

DOMESTIC  water  supply  systems  present  the  engineer  with  a  design 
problem  that  requires  combining  the  somewhat  empirical  rules  and 
formulae  in  use  with  the  more  or  less  exact  hydraulic  principles  involved. 
Unlike  heating  and  ventilating  layouts,  there  are  practically  no  definite 
data  for  estimating  the  quantity  of  water  likely  to  be  consumed  or  the 
probable  rate  of  water  flow  at  any  particular  moment. 

Metered  resujts  in  one  building  often  show  two  or  three  times  the 
metered  amount  in  another  building  of  the  same  size  and  with  the  same 
type  of  tenants.  In  hotels,  one  riser  will  often  have  an  almost  constant 
flow  that  may  never  be  reached  by  another  at  peak  load.  In  office 
buildings,  the  women's  toilets  show  a  far  greater  daily  consumption  than 
those  of  the  men,  yet  at  no  time  will  they  approach  the  hourly  consump- 
tion of  the  men's  toilet  during  the  first  hour  of  the  day.  This  condition 
has  led  to  a  multiplicity  of  rules  of  practice  which  vary  as  much  as  the 
data  used.  All  must  of  necessity  be  based  on  an  assumed  rate  of  con- 
sumption and  on  an  assumed  probability  of  simultaneous  use,  and  while 
the  formulae  employed  may  have  been  derived  on  sound  technical  bases 
the  assumptions  are  often  in  error. 

To  arrive  at  a  safe  standard,  the  approximate  rate  of  flow  of  each 
fixture  to  be  supplied  must  be  known  and  the  probable  number  of  fixtures 
in  use  at  any  one  time  must  be  assumed.  Obviously,  the  maximum 
number  of  fixtures  assumed  to  be  in  use  must  be  taken  at  the  peak  of 
demand  and  the  lines  must  be  made  adequate  to  supply  such  a  peak 
regardless  of  the  riser  or  branch  on  which  the  demand  may  occur.  This 
means  that  all  water  piping  under  the  usual  conditions  will  be  over-sized. 

In  tall  buildings  it  is  customary  to  divide  the  water  supply  systems, 
both  hot  and  cold,  into  sections  of  10  to  20  stories.  Such  zoning  or 
sectionalizing  is  for  the  purpose  of  avoiding  excessive  pressures  on  the 
fixtures  in  the  lower  stories  of  each  system.  This  limits  the  consideration 
of  water  pipe  sizes  to  horizontal  mains  and  to  risers  not  exceeding  20 
stories  in  height  or  about  200  ft1. 

lit  Is  impractical  to  attempt  to  size  piping  so  as  to  produce  the  proper  pressure  on  fixtures  at  different 
levels  by  employing  friction,  owing  to  the  fact  that  this  friction  will  be  built  up  to  the  amount  desired  only 
in  times  of  maximum  demand  and  at  all  other  times  the  friction  will  be  only  a  fraction  of  the  maximum 
friction  so  that  the  fixtures  by  this  method  are  subjected  to  a  varying  pressure  on  the  water  supply  line.  A 
much  more  practical  method  is  to  throttle  the  flow  at  the  fixture,  or  to  use  flo-w  regulators,  so  that  the 
quantity  of  water  delivered  will  approximate  the  fixture  demands  and  so  that  this  is  accomplished  without 
splashing  or  noise, 

599 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

For  the  purpose  of  this  chapter  the  following  terms  will  be  used  and 
should  be  clearly  distinguished  from  one  another: 

Maximum  Possible  Flow:  The  flow  which  would  occur  if  the  out- 
lets on  all  fixtures  were  opened  simultaneously.  This  condition  is  seldom, 
if  ever,  obtained  in  actual  practice  except  in  cases  of  gang  showers  con- 
trolled from  one  common  valve,  and  similar  conditions. 

Maximum  Probable  Flow:  The  maximum  flow  which  any  pipe  is 
likely  to  carry  under  the  peak  conditions.  This  is  the  most  important 
amount  to  be  considered  in  pipe  sizing. 

Average  Probable  Flow:  The  flow  likely  to  be  required  through  the 
line  under  normal  conditions. 

It  is  evident  that  any  pipe  adequate  to  take  care  of  the  maximum 


Maximum  Probable  Percentage  of  Use 

*—  N>  u>  -t*  en  a*  ^  oo  UD  c 
oooo  ooo^ooc 

y 

\ 

\ 

N 

\ 

\ 

\ 

s, 

ss 

\ 

""-^ 

n 

•*»^ 

ia 

^ 

s. 

--y. 

2j 

iures_^ 

11^. 

—    - 

—  . 

X 

"*• 

""•i 

^t 

'3/1 

rfi 

xtu 

m 

1< 

?J 

fr-tf- 

,2% 

-44- 

^ 

^c 

I            2       345       8  10          20    30  40  50     80  100       200            500       1000 

Number  of  Fixtures 

FIG.  1.    CHART  SHOWING  RELATION  BETWEEN  NUMBER  OF  FIXTURES  AND 
MAXIMUM  PROBABLE  PERCENTAGE  OF  USE 


probable  flow  will  also  be  more  than  able  to  take  care  of  the  average 
probable  flow,  and  hence  the  latter  has  no  bearing  on  the  pipe  size, 

MAXIMUM  PROBABLE  FLOW 

There  are  two  factors  to  be  considered  in  calculating  the  maximum 
probable  flow,  namely,  (1)  the  quantity  of  water  that  will  flow  from  the 
outlets  when  they  are  open,  and  (2)  the  number  of  outlets  likely  to  be  open 
at  the  same  time.  Table  1  shows  the  maximum  approximate  rate  of  flow 
from  each  fixture  when  it  is  in  use,  and  will  serve  as  a  guide  in  estimating 
maximum  probable  flow  demands  although  there  is  considerable  variation 
in  different  fixtures  and  valves.  Probably  the  flow  under  normal  water 
pressures,  or  with  the  pressure  properly  throttled,  will  not  differ  greatly 
from  the  values  stated.  With  the  aid  of  this  table  it  is  possible  to  calculate 
the  maximum  possible,  flow  with  all  outlets  open  in  both  the  hot  and  cold 
water  lines. 

600 


CHAPTER  35  — 


SUPPLY  PIPING 


Factor  of  Usage 

To  obtain  the  maximum  probable  flow  it  is  necessary  to  multiply  the 
maximum  possible  flow  by  a  factor  of  usage,  and  this  factor  varies  with  the 
installation  and  the  number  of  fixtures  in  the  installation.  It  is  evident 
that  with  two  fixtures  it  is  quite  possible  that  both  will  at  some  time  be  in 
operation  simultaneously.  With  200  fixtures,  it  is  unlikely  the  entire 
200  would  ever  operate  at  the  same  time.  Consequently,  the  factor  of 
usage  reduces  as  the  number  of  fixtures  becomes  greater,  all  other  things 
being  equal.  On  the  other  hand  it  is  probable  that  outside  of  flush  valve 

TABLE  1.    APPROXIMATE  FLOW  FROM  FIXTURES  UNDER  NORMAL  WATER  PRESSURES 


FIXTURES 

COLD  WATER 
(GALLONS  PER 
MINUTE) 

HOT  WATER 
(GALLONS  PER 
MINUTE) 

Watsr-closets,  flush  valve                            .          

50  a 

0 

Water-closets,  flush  tank    

18 

0 

Urinals,  flush  valve 

40  a 

0 

Urinals,  flush  tank  

18 

0 

Urinals,  automatic  tank                 .     .                       

1 

0 

Urinals  perforated  pipe  per  foot 

10 

0 

Lavatories 

3 

3 

Showers,  5  to  6  J^  in.  heads  

3 

3 

Showers,  tubular                                  *    .            ...  . 

6 

6 

Needle  bath  

30 

30 

Shampoo  sorav                                            

1 

1 

Liver  spray 

2 

2 

Manicure  table                                             .            

1H 

1H 

Baths,  tub  

5 

5 

Kitchen  sink                                     .            

4 

4 

Pantry  sink,  ordinary 

2 

2 

Pantry  sink,  large  bibb                                     

6 

6 

Slop  sinks     

4 

4 

Wash  travs                        .               ..               

3 

3 

aActual  tests  on  water-closet  flush  valves  indicate  40  gpm  as  the  maximum  rate  of  flow  with  30  Ib  pres- 
sure at  the  valve;  this  would  increase  to  60  gpm  (about  50  per  cent)  at  90  Ib  pressure.  The  50  gpm  has  been 
taken  as  an  average  flow;  possibly,  with  very  low  pressures  just  sufficient  to  operate  the  flush  valve,  30  gpm 
could  be  allowed  with  safety.  Urinal  flush  valves  would  vary  proportionately  in  the  same  manner. 

fixtures,  the  factor  of  usage  would  never  be  less  than  about  25  per  cent  no 
matter  how  many  fixtures  were  installed,  provided  no  fixtures  in  excess  of 
those  required  for  the  actual  occupancy  were  included. 

This  factor,  beginning  at  100  per  cent  for  two  ordinary  fixtures, 
decreases  rapidly  until  5  fixtures  are  reached  and  then  becomes  almost 
constant,  as  shown  in  the  upper  curve,  Fig.  1.  This  applies  to  a  normal 
building  and  not  to  institutions  where  the  inmates  may  all  be  required, 
for  instance,  to  bathe  on  certain  days  of  the  week  and  at  certain  hours  of 
those  days.  In  such  special  cases  a  new  factor  of  usage  must  be  developed 
based  on  the  maximum  probable  usage  under  the  conditions  ^involved. 
For  flush  valve  fixtures  the  quantity  of  water  is  greater,  but  owing^to  the 
short  duration  of  the  flush,  the  simultaneous  usage  drops  more  rapidly  sc 
as  to  reach  1  per  cent  for  1000  fixtures  as  shown,  on  lower  curve,  Fig.  I2, 

*This  can  be  proved  by  assuming,  for  example,  1000  water-closets  which  would  not  be  used  more  that 
six  times  per  hour  (or  once  every  10  minutes)  and  which  require  from  5  to  7  gal  per  flush  or  an  average  o 
about  6  gal.  If  these  closets  were  all  being  used  at  their  utmost  capacity,  the  water  demand  would  b< 
600  gpm.  But  average  use  would  be  about  one-third  of  this  and  peak  conditions  would  be  in  the  neigh 
borhood  of  twice  the  average,  or  about  400  gpm  as  the  maximum  that  would  ever  develop.  Assuming  5( 
gpm  as  the  maximum  rate  of  flow  per  closet  and  1  per  cent  of  the  total  closets  in  operation,  the  rate  wouk 
be  50  gpm  X  1  per  cent  of  1000  or  500  gpm,  This  is  100  gpm  higher  than  obtained  by  the  first  method 
Indicating  an  additional  factor  of  safety  over  the  first  method. 

601 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

Example  1 .  Assume  that  in  a  normal  building,  such  as  a  residential  hotel  or  an  apart- 
ment house,  there  are  50  flush  valve  water-closets,  50  lavatories,  50  sinks  and  50  baths, 
and  that  it  is  desired  to  determine  the  maximum  probable  flow  in  a  line  supplying  all  of 
these  fixtures  with  both  hot  and  cold  water.  Fig.  1  shows  a  maximum  probable  use  for 
50  water  closets  of  about  8  per  cent  and  for  150  ordinary  fixtures,  of  about  31  per  cent. 
Therefore: 

Cold  Water 

50  W.  C,  x  50  gpm  at  8  per  cent 200  gpm 

50  Lavs,  x  3  gpm 150  gpm 

50  Sinks  x  4  gpm 200  gpm 

50  Baths  x  5  gpm 250  gpm 

150  Fixtures 600  gpm  at  31  per  cent  186  gpm 


Total  maximum  probable  flow  of  cold  water 386  gpm 

Hot  Water 

50  W.  C , None 

50  Lavs,  x  3  gpm _ 150  gpm 

50  Sinks  x  4  gpm 200  gpm 

50  Baths  x  5  gpm 250  gpm 

150  Fixtures 600  gpm  at  31  per  cent  186  gpm 


Total  for  main  supplying  cold  and  hot  water 572  gpm 

It  should  be  noted  that  this  is  a  rate  of  flow  or  an  instantaneous  demand. 

KIND  OF  PIPE  USED 

Before  entering  into  the  actual  sizing  of  pipe,  it  is  necessary  to  consider 
the  kind  of  pipe  to  be  used,  and  to  make  suitable  allowance  for  corrosion 
and  fouling  during  the  lifetime  of  the  system.  For  example,  if  brass, 
copper  or  alloy  pipe  is  contemplated,  it  is  probable  that  the  quantities 
indicated  in  Example  1  are  ample;  if  galvanized  pipe  is  to  be  used,  then  it 
is  quite  likely  that  after  a  period  of  say  15  years  the  area  may  be  decreased 
as  much  as  25  per  cent  and  the  quantitities  of  water  assumed  should  be 
increased  by  35  per  cent  to  allow  for  this  reduction  of  area;  if  the  water 
contains  lime  it  is  possible  that  50  per  cent  of  the  area  may  be  lost  and  in 
such  cases  the  flow  should  be  doubled  and  no  branch  pipe  connected  to 
fixtures  should  be  less  than  %  in.  In  all  of  the  following  calculations,  the 
assumption  is  made  that  the  water  is  fairly  good  and  that  a  corrosion 
resistant  type  of  pipe  is  to  be  used. 

SIZING  A  DOWN-FEED  RISER 

Down-feed  systems  are  commonly  used  for  tall  buildings.  In  sizing  a 
riser  arranged  for  down-feed,  the  gravity  head  permits  a  pressure  drop 
that  is  almost  prohibitive  in  an  up-feed  risen  There  is  a  gain  in  riser  head 
of  0.43  X  100  or  43  Ib  per  100  ft  of  run  and  hence  it  is  quite  permissible 
to  size  such  a  riser  on  the  basis  of  a  pressure  drop  of  30  Ib  per  100  ft  of  run, 
as  the  difference  between  the  43  Ib  generated  and  the  30-lb  drop  under 
maximum  probable  demand  is  ample  to  take  care  of  the  friction  caused  by 

602 


CHAPTER  35 — WATER  SUPPLY  PIPING 


the  fittings.  This  method  applied  to  the  typical  riser  shown  in  Fig.  2 
gives  the  schedule  of  sizes  indicated  in  Table  2  for  any  flow  from  5  to  250 
gal. 


CO   CO    CO   CO    CO   CO 


CO    CO   CO    CO    CO 


COCO    CO    COCO    COCO    COCO 


CO         CO         CO 


COCO         CO         COCO         COCO        COCN 


CO         CO     -   CO        CO        CO 


CO         CO         CO 


CO         COCOCO         tN       CO         CO       CN         CO       CO 


CO         CO       CO         CO       (N         CO       CO 


CO         CO        CO         CO       CO 


COCO         CO         COCO         COCO        COCO 


I          iH          iH 

Pi"' 

hcoftsOBsQ^^^^^^^^k* 

tf 

go 
«s 

*g 

cccccccifcEc^^^^^^EE^jS 
§fg|Sg**i§g§|£|£:B£SSl3o 

t't't  tttttttt  tut  mitEi 

6n          '  '^  '  'ft}  "O1  '  '  P-  '  '  O       ^ighj^^^&JO^^QOfiQ-^  Q 

(NtN 

Is 

SIZING  AN  UP-FEED  RISER 

When  the  riser  is  an  up-feed,  the  opposite  condition  occurs;  that  is, 
there  is  a  drop  in  pressure  as  the  top  of  the  riser  is  approached,  due  to  the 
natural  reduction  in  the  gravity  pressure,  and  to  this  must  be  added  the 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


pipe  friction  plus  that  introduced  by  the  pipe  fittings,  all  of  which  produce 
an  excessive  drop  when  compared  to  the  conditions  existing  with  a  down- 
feed  riser. 

To  size  an  up-feed  riser  the  minimum  pressure  of  the  street  main,  or 
other  source  of  supply,  should  be  ascertained  and  from  this  should  be 
subtracted  the  pressure  to  be  maintained  at  the  highest  fixture,  namely, 
15  Ib  per  square  inch,  plus  the  height  in  feet  above  the  source  of  water 
pressure,  multiplied  by  0.43  to  change  from  feet  of  head  to  pounds  of 
pressure.  The  total  length  of  run  from  the  source  of  pressure  to  the 
farthest  and  highest  fixture  should  be  ascertained,  and  this  should  be 
changed  to  equivalent  length  of  run  to  allow  for  the  loss  occasioned  by 
the  pipe  fittings.  Table  3  gives  the  additional  lengths  necessary  to  allow 
for  the  various  fittings  and  valves.  The  drop  allowable  in  pressure  per 
100  ft  of  run  may  then  be  obtained  by  multiplying  the  surplus  pressure 
(over  that  required  for  the  gravity  head  and  to  supply  15  Ib  at  the  fixture) 
by  100  and  by  dividing  this  by  the  equivalent  length  of  run  to  the  farthest 
or  highest  fixture. 

Example  8.  Assume  a  street  pressure  of  60  Ib,  the  height  of  the  highest  fixture  50  ft, 
and  the  length  of  the  longest  run  200  ft.  Without  knowing  the  additional  length  of  pipe 
to  be  added  for  the  fittings  it  will  be  assumed  that  this  is  about  100  ft.  The  surplus 
pressure  which  will  be  available  for  pressure  drop  will  then  be 

60  Ib  -  (15  Ib  +  50  ft  X  0.43  Ib)  = 

60  Ib  -  (15  Ib  +  21.5  Ib)  =  23.5  Ib 


To  change  this  into  drop  per  100  ft: 

23.5  Ib  X  100 


200  ft  +  100  ft 


7.8  Ib  per  100  ft. 


The  pipe  may  then  be  sized  from  the  maximum  probable  flow  by  selecting  a  size  that 
does  not  give  a  drop  in  excess  of  7.8  Ib  per  100  ft. 

It  will  be  seen  from  Example  2  that  it  is  impossible  to  size  up-feed 
risers  without  determining  the  drop  allowable  in  both  the  horizontal  feed 
mains  and  the  toilet  room  branches.  Having  once  ascertained  this  allow- 
able drop,  it  is  simply  a  matter  of  applying  it  throughout  the  system. 


TABLE  3. 


APPROXIMATE  ALLOWANCES  FOR  FITTINGS  AND  VALVES 
IN  FEET  OF  STRAIGHT  PIPE 


TTPE  or  FITTING  OR  VALVE 


SIZE  OF  PIPE 
(INCHES) 

90-Deg 
Elbow 

45-Deg 
Elbow 

Return 
Bend 

Gate 
Valve 

Globe 
Valve 

Angle 
Vafve 

K 

4 

3 

8 

2 

48 

8 

H 

5 

3 

10 

3 

60 

10 

l 

5 

3 

10 

3 

60 

10 

IK 

6 

4 

12 

3 

72 

12 

IX 

7 

5 

14 

4 

84 

14 

2 

7 

5 

14 

4 

84 

14 

m 

10 

7 

20 

5 

120 

20 

3 

12 

8 

24 

6 

144 

24 

4 

18 

13 

36 

9 

216 

36 

5 

25 

18 

50 

13 

300 

50 

6 

30 

21 

60 

15 

360 

60 

604 


CHAPTER  35 — WATER  SUPPLY  PIPING 


HORIZONTAL  SUPPLY  MAINS 

The  horizontal  mains  supplying  the  risers  at  the  top  of  a  down-feed 
system  must  be  liberally  sized  unless  the  house  tank  is  set  at  a  much 
higher  elevation  than  usual.  To  provide  a  gravity  head  on  the  highest 
fixtures  of  15  Ib  per  square  inch  it  is  necessary  for  the  water  line  in  the 
house  tank  to  be  nearly  40  ft  higher,  and  with  the  line  loss  considered  this 
becomes  about  45  ft.  Such  heights  are  not  often  practical  and  as  a  result 
the  pressure  on  the  highest  fixtures  either  is  reduced  to  7  Ib  (which  is 
sufficient  to  operate  a  flush  valve),  or  flush  tank  water-closets  are  sub- 
stituted, or  a  separate  cold  and  hot  water  supply  is  installed  with  a  small 
pneumatic  tank  to  give  the  increase  in  pressure  necessary.  The  chief 
objection  to  the  use  of  a  pneumatic  tank  is  that  a  separate  hot  water 
heater  is  required  and  this  heater  must  be  located  either  sufficiently 
below  the  highest  fixtures  to  obtain  a  gravity  circulation,  or  it  must  be 
provided  with  a  circulating  pump  in  order  to  force  the  hot  water  to  the 
top  floor  level. 

The  most  common  solution  is  to  place  the  house  tank  as  high  as  the 
structural  and  architectural  conditions  will  permit  and  then  to  use 
liberally-sized  lines  between  the  house  tank  and  the  upper  fixtures,  say  for 
the  two  top  stories,  below  which  the  riser  sizes  may  be  reduced  to  those 
indicated  in  Fig.  2  and  Table  2.  Where  the  house  tank  is  only  one  story 
above  the  top  fixtures,  flush  tank  water-closets  must  be  used  and  the 
drop  in  the  entire  run  from  the  house  tank  down  to  the  farthest  fixture 
should  not  exceed  1  Ib ;  the  less,  the  better.  This  means  that  if  the  total 
equivalent  run  to  the  farthest  top  fixtures  supplied  is  300  ft,  the  drop  per 

100  ft  should  not  exceed  1  lb  *  1Q°  or  0.33  Ib  per  100  ft.     The  friction 

oUU 

curves  shown  in  Fig.  3  may  be  used  for  quickly  determining  the  proper 
size  of  pipe  to  give  any  desired  drop  in  pounds  per  100  ft  of  equivalent  run. 

OVERHEAD  DISTRIBUTION  MAIN 

Example  3.  Suppose  an  installation  has  a  house  tank  in  which  the  water  line  is  20  ft 
above  the  level  of  the  top  fixtures  to  be  supplied  and  that  the  length  of  run  to  the 
farthest  fixtures  on  this  level  is  400  ft  with  the  pipe  fittings  adding  another  200  ft, 
making  an  equivalent  length  of  600  ft.  What  would  be  the  size  of  main  coming  out 
of  the  tank  where  a  maximum  flow  rate  of  400  gpm  may  be  expected,  of  the  horizontal 
main  where  a  maximum  flow  rate  of  200  gpm  may  be  expected,  and  of  the  riser  down  to 
the  fixture  level  where  the  maximum  flow  rate  is  approximately  100  gpm? 

Here  the  level  of  the  water  in  the  house  tank  is  20  ft  above  the  faucet  of  the  highest 
fixture  and  the  gravity  pressure  will  be 

0.43  Ib  X  20  ft  «  8.6  Ib 

and,  if  a  total  pressure  drop  of  1  Ib  is  assumed,  the  pressure  on  the  farthest  fixture  under 
times  of  peak  load  will  be 

8.6  Ib  -  1  Ib  -  7.6  Ib 

while  the  drop  per  100  ft  of  equivalent  run  will  have  to  be 

Hb  X  100 


600 


«  0,1667  Ib. 


Referring  to  Fig.  3  it'  will  be  noted  that  where  the  flow  through  the  main  is  400  gpm,  ^an 
8  in.  pipe  would  be  required;  that  where  the  flow  is  reduced  to  200  gpm,  a  6-m.  pipe 

605 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


would  be  sufficient ;  and  that  where  the  flow  is  100  gpm  in  the  riser  branch  and  riser,  a 
5-in.  size  would  be  correct.  Of  course  these  are  somewhat  excessive  flows  and  the  head 
from  the  tank  is  small  so  that  large  sizes  are  to  be  expected.  It  would  be  necessary  to 
carry  a  5-in.  riser  down  to  the  branch  to  the  top  floor,  then  reduce  to  4  in.  for  the 
branch  to  the  floor  below  the  top,'  and  below  this  the  sizes  in  Table  2  could  be  followed. 
In  such  a  case,  flush  tank  closets  should  doubtless  be  substituted. 

Had  the  tank  been  set  10  ft  higher,  the  head  available  to  be  used  up  in  friction,  but 


1000 
900 
800 
700 
600 
500 
400 
a.  300 


" 


100 


100 
90 

80 
70 
60 
50 
40 


30 


20 


10 


7 


7\ 


For  Clean 


Iron 


Pipe 


VO.T  0.2     0.3  0.4  0.5  1.0  2.0     3.0  4.0  5.0  10  20      30   40  50  100 

Friction  Loss  per  100  Ft  Straight  Pipe  in  Pounds  per  Square  Inch 

Fro.  3,    CHART  GIVING  FRICTION  LOSSES  FOR  VARIOUS  RATES  OF  FLOW  OF  WATER 

still  giving  th^same  pressure  at  the  top  fixtures,  would  have  been  0.43  Ib  X  10  ft  or  4.3  Ib 
greater  and  this,  with  the  1  Ib  drop  used  previously,  would  give  a  total  allowable  drop  of 

1  Ib  -f  4.3  Ib  =  5.3  Ib 

which,  divided  by  the  600  ft  equivalent  run  gives  a  drop  per  100  ft  of 

5.3  X  100 


600 


»  0.9  Ib 


arid,  with  this  drop,  the  sizes  according  to  the  chart  (Fig.  3)  are  6  in.,  4  in.,  and  4  in,, 

606 


CHAPTER  35 — WATER  SUPPLY  PIPING 


respectively,  while  if  the  run  is  reduced  to  200  ft  instead  of  600  ft,  the  allowable  drop  will 
be 


5.3  Ib  X  100 
200 


=  2.7  Ib  per  100  ft. 


This  gives  5  in.,  4  in.,  and  3  in.,  respectively,  for  the  flows  of  400,  200,  and  100  gpm. 


Water  Line 

n 

h       

House  Supply 

„• 

^ 

Fire  Reserve 

House  Tank-!"' 

* 

•-              "    -— 

j 

r 

~2 

* 

5" 

'                   199               4"                  ' 

296 

5" 

297 

199 

4" 

4W.C.-F.V. 
2U.-F.V.                            202 
3  Lav. 

•4" 

6W.C.-F.V 
4  Lav. 

137 

3" 

IS.  S. 

8th. 

4->- 

+-V- 

179 

2*' 

4W.C.-F.V. 
2U.-F.V.                            189 
3  Lav. 

«f 

6W.C.-F.V. 
4  Lav. 

136 

2f 

1S.S. 

7th.         : 

h-  >• 

1—  >• 

I1  —  >• 

165 

2" 

4W.  C.-F.V. 
2U.-F.V.                            182 
3  Lav. 

2" 

6W.C.-F.V. 
4  Lav. 

134 

2" 

1S.S. 

6th. 

*—  ^- 

t—^- 

*—  *• 

134 

2" 

4W.C.-F.V. 
2U.-F.V.                           173 
3  Lav. 

2" 

6W.  C.-F.V. 
4  Lav. 

133 

2" 

1S.S. 

5th. 

•  ^ 

I    3^ 

4W.C.-F.V. 

3W.C.-F.V. 

19 

1" 

10  Lav.                               164 

2" 

2U.-F.V. 

131 

2" 

ILav. 

3  Lav. 

1S.S. 

4th.         J 

*—  ^ 

:«—  ** 

*-*• 

10 

r 

1S.S.                                 149 

2" 

4W.  C.-F.V. 
2U.-F.V. 

129 

2" 

2  Lav. 

3  Lav.     ' 

3rd.      . 

>—  >• 

*-*- 

8 

i" 

1  S,  S.                                  98 

4' 

2W.C.-F.V. 
1U.-F.V. 
ILav, 

127 

2" 

3W.C.-F.V. 
ILav. 

2nd. 

it—  >• 

i—  *• 

,i  > 

4 

4 

!.—  a^ 

1  S.  S.,                                  50 

«—  *^ 

1WC.-F.V. 

4 

r 

i  -,w 

1S.S. 

1st 


(1\  (2)  (3) 

FIG.  4,    TYPICAL  LAYOUT  FOR  DOWN-FEED  SYSTEM 

From  Example  3  it  is  evident  that,  while  the  down-feed  system  possesses 
certain  economies  in  size  for  the  riser  portion,  it  is  quite  likely  to  involve 
large  distribution  main  sizes,  especially  when  the  tank  is  not  elevated  to  a 
considerable  degree. 

SIZING  A  PIPING  SYSTEM 

Example  4>  Fig.  4  shows  a  typical  layout  with  three  risers  extending  eight  stories  arid 
with  the  fixtures  noted  on  each  floor.  First  this  will  be  solved  for  a  down-feed  arrange- 
ment assuming  that  the  level  of  the  water  in  the  house  tank  is  30  ft  above  the  fixtures  on 

607 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


the  top  floor,  that  the  length  of  run  from  the  tank  to  the  farthest  fixture  is  200  ft,  equiva- 
lent length  of  fittings  100  ft,  and  the  pressure  required  at  the  fixture  is  7  Ib. 

The  30-ft  head  is  equal  to  a  static  pressure  of  0.43  X  30  or  12.9  Ib  per  square  inch  and 
to  maintain  a  pressure  of  7  Ib  at  the  highest  fixtures  the  drop  allowable  in  pressure  is 
12.9  -  7.0  Ib  or  5.9  Ib.  As  the  total  equivalent  run  is  300  ft,  this  is  a  drop  per  100  ft  of 
1.97  Ib,  or  practically  2  Ib.  Therefore,  all  risers  and  mains  from  the  top  floor  back  to  the 

TABLE  4.    TYPICAL  CALCULATION  OF  PIPE  SIZES  ON  DOWN-FEED  RISER  WITH 
FLUSH  VALVE  WATER-CLOSETS  AND  URINALS 

Riser  No.  1.    (See  Fig.  4) 


FLOOR 

OF 

BLDG. 

FIXTURES 

ON 

FLOOR 

GPM 

PER 

FIXTURE 

MAXIMUM 
FIXTURES 
GPM 

PROBABLE 
USE 
(PER  CENT) 

PROBABLE 
FIXTURES 
GPM 

PROBABLE 
RISER 
GPM 

ALLOWABLE 
DROP 
LB  PER  100  FT 

PIPE 
SIZE 
IN. 

1st 

1  S.  S. 

4 

4 

100 

4 

4 

30 

H 

2nd 

1  S.  S. 

4 

4 

100 

4 

8 

30 

% 

3rd 

1  S.  S. 

4 

4 

3 

12 

80 

10 

10 

30 

H 

4th 

10  Lav. 

3 

30 

13 

42 

45 

19 

19 

30 

I 

5th 

4  W.  C. 
2U. 

50 
40 

200 
80 

6 

3  Lav. 

3 

280 
9 

40 

112 

16 

51 

45 

22 

134 

30 

2 

6th 

4  W.  C. 
2  U. 

50 
40 

200 
80 

12 
3  Lav. 

3 

560 
9 

25 

140 

19 

60 

42 

25 

165 

30 

2 

7th 

4  W.  C. 

2U. 

50 
40 

200 
80 

18 
3  Lav. 

3 

840 
9 

18 

151 

22 

69 

40 

28 

179 

30 

2M 

8th 

4  W.  C. 
2U. 

50 
40 

200 
80 

24 
3  Lav. 

3 

1120 
9 

15 

168 

25 

78 

40 

31 

199 

2 

4 

tank  must  be  sized  on  the  basis  of  a  drop  of  2  Ib  per  100  ft.  Tables  4,5,6  and  7  show  the 
schedule  for  Risers  No.  1,  2  and  3  with  the  maximum  possible  flow  taken  from  Table  1, 
the  percentage  of  use  at  the  peak  taken  from  Fig.  1,  and  the  maximum  probable  flow  at 
the  peak  worked  out  for  each  portion  of  the  riser,  the  riser  sizes  being  taken  from  Table  2 
as  far  as  possible  and  from  Fig.  3  where  the  amounts  exceed  the  values  given  in  this 
table;  a  drop  of  30  Ib  per  100  ft  is  used  except  on  the  riser  from  the  top  floor  back  to  the 
tank  where  2  Ib  per  100  ft  is  the  allowable  limit. 

The  reduction  in  pipe  size  which  would  occur  if  flush  tank  water-closets  were  used  on 
the  top  floor  and  only  3  Ib  pressure  used  on  the  fixtures  is  given  in  Tables  8  and  9. 

608 


CHAPTER  35 — WATER  SUPPLY  PIPING 


This  illustrates  why  flush  tank  closets  so  frequently  are  substituted  on  the  uppermost 
floor  when  a  house  tank  is  the  source  of  water  pressure. 

If  it  is  now  assumed  that  Riser  No.  1  is  to  be  fed  from  the  bottom  and  the  minimum 
street  pressure  is  75  Ib  with  the  top  fixture  of  the  riser  80  ft  above  the  main,  the  problem 

TABLE  5.    TYPICAL  CALCULATION  OF  PIPE  SIZES  ON  DOWN-FEED  RISER  WITH 
FLUSH  VALVE  WATER-CLOSETS  AND  URINALS 

Riser  No.  2.    (See  Fig.  4) 


FLOOR 

OF 

BLDG. 

FIXTURES 

ON 

FLOOR 

GPM 

PER 

FIXTURE 

MAXIMUM 
FIXTURES 
GPM 

PROBABLE 
USE 
(PER  CENT) 

PROBABLE 
FIXTURES 
GPM 

PROBABLE 
RISER 
GPM 

ALLOWABLE 
DROP 
LEPER  100  FT 

PIPE 
SIZE 

IN. 

1st 

1  W.  C. 

50 

50 

100 

50 

50 

30 

IK 

2nd 

2  W.  C. 

1  U. 

50 
40 

100 
40 

4 
1  Lav. 

3 

190 
3 

50 
100 

95 
3 

98 

30 

1H 

3rd 

4  W.  C. 
2  U. 

50 
40 

200 
80 

10 
3  Lav. 

3 

470 
9 

30 

141 

4 

12 

70 

8 

149 

30 

2 

4th 

4  W.  C. 
2  U. 

50 
40 

200 
80 

16 
3  Lav. 

3 

750 
9 

20 

150 

7 

21 

70 

14 

164 

30 

2 

5th 

6  W.  C. 

50 

300 

22 
4  Lav. 

3 

1050 
12 

15 

157 

11 

33 

48 

16 

173 

30 

2 

6th 

6  W.  C. 

50 

300 

28 
4  Lav. 

3 

1350 
12 

12 

162 

15 

45 

45 

20 

182 

30 

2 

7th 

6  W.  C. 

50 

300 

34 
4  Lav. 

3 

1650 
12 

10 

165 

19 

57 

42 

24 

189 

30 

2H 

8th 

0  W.  C. 

50 

300 

40 
4  Lav. 

3 

1950 
12 

9 

175 

23 

69 

40 

27 

202 

2 

4 

would  be  solved  by  determining  the  maximum  rate  of  flow  in  each  portion  of  the  riser  as 
shown  in  Table  10  and  then  finding  the  allowable  drop  which  can  be  used  per  100  ft. 
The  80  ft  of  riser  height  will  use  up 


0.43  Ib  X  80 
609 


34.4  Ib 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


and  the  pressure  at  the  top  of  the  required  15  Ib  will  make  the  total  reduction  49.4  Ib, 
leaving  a  balance  of  25.6  Ib  which  may  be  used  up  in  friction.    If  the  distance  from  the 

TABLE  6.    TYPICAL  CALCULATION  OF  PIPE  SIZES  ON  DOWN-FEED  RISER  WITH 

FLUSH  VALVE  WATER-CLOSETS  AND  URINALS 

Riser  No.  3.    (See  Fig.  4) 


FLOOR 

OF 

BLDG. 

FIXTURES 

ON 

FLOOR 

GPM 

PER 

FIXTURE 

MAXIMUM 
FIXTURES 
GPM 

PROBABLE 
USE 
(PER  CENT) 

PROBABLE 
FIXTURES 

GPM 

PROBABLE 
RISER 
GPM 

ALLOWABLE 
DROP 
LB  PER  100  FT 

PIPE 
SIZE 

IN. 

1st 

IS.  S. 

4 

4 

100 

4 

4 

30 

X 

2nd 

3W.  C. 

1  Lav. 

50 
3 

150 
3 

80 

120 

2 

7 

100 

7 

127 

30 

2 

3rd 

OW.  C. 

0 

000 

3 

2  Lav. 

3 

150 
6 

80 

120 

4 

13 

70 

9 

129 

30 

2 

4th 

3  W.  C. 

50 

150 

6 

ILav. 
1  S.  S. 

3 

4 

300 
3 

4 

40 

120 

6 

20 

55 

11 

131 

30 

2 

5th 

0  W.  C. 

0 

000 

6 
1  S.  S. 

4 

300 

4 

40 

120 

7 

24 

53 

13 

133 

30 

2 

6th 

0  W.  C. 

0 

000 

6 
1  S,  S. 

4 

300 
4 

40 

120 

8 

28 

51 

14 

134 

30 

2 

7th 

0  W.  C. 

0 

000 

6 
1  S.  S. 

4 

300 
4 

40 

120 

9 

32 

50 

16 

136 

30 

2H 

8th 

0  W.  C. 

0 

000 

6 
IS.  S. 

4 

300 
4 

40 

120 

10 

36 

48 

17 

137 

2 

3 

street  main  to  the  bottom  of  the  riser,  which  will  be  assumed  to  be  the  farthest  one  on  the 
horizontal  line,  is  100  ft,  and  if  the  fittings  are  sufficient  to  add  another  100  ft,  as  well  as 
the  80  ft  of  vertical  distance  up  the  riser,  the  total  equivalent  run  will  be  280  ft,  which 
will  be  taken  as  an  even  300  ft.  Then  the  allowable  drop  per  100  ft  will  be 


25.6  Ib  X  100 
300 


•  8.5  Ib 


and  the  sizes  shown  in  Fig.  5  are  based  on  this  amount  of  drop.    Of  course  the  other 

610 


CHAPTER  35 — WATER  SUPPLY  PIPING 


risers  will  have  the  same  maximum  flows  at  the  bottom  as  they  formerly  had  at  the  top, 
namely  202  and  137  gal,  respectively,  for  Risers  No.  2  and  3.  Combining  these  maxi- 
mum flows  in  the  same  manner  as  pursued  in  the  down-feed  system  it  is  seen  that  the 
maximum  flow  between  Riser  No.  2  and  Riser  No.  3  is  296  gpm,  and  between  Riser 
No.  3  and  the  street  main,  297  gpm  which  at  a  drop  of  8.5  Ib  gives  the  main  sizes 
indicated.  It  will  be  noted  that  in  determining  the  maximum  flow  in  an  up-feed  riser 

TABLE  7.    SIZE  OF  DISTRIBUTION  MAIN  FOR  DOWN-FEED  SYSTEMS  (SEE  FIG.  4) 


RISER 
No. 

FIXTURES 

GPM  PER 

FIXTURE 

MAXIMUM 
FIXTURES 

GPM 

PROBABLE 
USE 
(PER  CENT) 

PROBABLE 

GPM 

ALLOWABLE 
DROP  (Ls 
PER  100  FT) 

SIZE  OF 

MAIN 
(INCHES) 

16  W.  C. 

8U. 

50 
40 

800 
320 

1 

24 
22  Lav. 
3  S.  S. 

3 

4 

1120 
66 
12 

15 

168 

25 

78 

40 

31 
199 

2 

4 

35  W.  C. 
5U. 

50 
40 

1750 
200 

2 

64 
23  Lav. 

3 

3070 
69 

8 

245 

48 

147 

35 

51 
296 

2 

5 

6  W.  C. 

50 

300 

3 

70 
4  Lav. 
6S.  S. 

3 
4 

3370 
12 

24 

7 

236 

58 

183 

33 

61 
297 

2 

5 

it  is  necessary  to  begin  at  the  top  floor  and  work  down  instead  of  beginning  at  the  bottom 
floor  and  working  up  as  was  done  in  the  down-feed  sizing. 

SIZING  UP-FEED  AND  DOWN-FEED  HOT  WATER  SYSTEMS 

Hot  water  supply  systems,  when  of  the  circulating  type,  have  a  few 
differences  to  be  considered  although  the  same  general  principles  of  sizing 
apply  to  these  lines  as  to  the  cold  water  lines.  Owing  to  the  fact  that 
there  are  no  flush  valves  on  the  hot  water  piping  and  also  because  many 
plumbing  fixtures  have  no  hot  water  connections,  the  sizes  of  the  hot 
water  piping  in  general  will  be  considerably  less  than  the  cold  water 
piping  in  the  same  building.  On  the  other  hand  it  is  almost  invariably 
required  that  a  gravity  circulation  be  kept  up  in  such  hot  water  lines  and 
this  often  has  a  considerable  influence  on  the  size.  There  are  three 
methods  of  arranging  circulation  lines,  as  follow ; 

1.  By  using  the  plain  up-feed  with  a  return  carried  back  from  the  top  of  the  riser  and 
paralleling  it, 

2.  By  carrying  a  supply  riser  up  in  one  location  thus  supplying  fixtures  on  up-feed, 
then  crossing  over  at  the  top  and  coming  down  past  another  collection  of  fixtures  and 
supplying  these  by  a  down-feed. 

3.  By  carrying  all  of  the  water  to  the  top  of  the  building  and  dropping  risers  wherever 
needed,  feeding  all  hot  water  on  a  down-feed  system* 

611 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


TABLE  8.    TYPICAL  CALCULATION  OF  PIPE  SIZES  ON  DOWN-FEED  RISERS  WITH 
FLUSH  TANK  WATER-CLOSETS  AND  URINALS  ON  TOP  FLOOR  ONLY  (SEE  FIG.  4) 


FLOOR 

FIXTURES 

GPM 

MAX. 

PROBABLE 

PROBABLE 

PROBABLE 

ALLOWABLE 

PIPE 

OF 

ON 

•PER 

FlXT. 

USE 

USE 

RISER 

DROP 

SIZE 

BLDG. 

FLOOR 

FlXT. 

GPM 

(PER  CENT) 

GPM 

GPM 

LB  PER  100  FT 

IN. 

Riser  No.  1 


7th  and 
below 

12  W.  C. 

6U. 

50 
40 

600 
240 

18 
19  Lav. 
3S.  S. 

3 

4 

840 
57 
12 

IS 

151 

22 

69 

40 

28 

179 

30 

2^ 

8th 

OW.  C. 
OU. 

00 
00 

000 
000 

18 
4  W.  C. 
2U. 
3  Lav. 

18 
IS 
3 

840 

72 
36 
9 

18 

151 

31 

186 

37 

69 

220 

3.3 

4 

Riser  No. 


7th  and 

29  W.  C. 

50 

1450 

below 

5U. 

40 

200 

34 

1650   - 

10 

165 

19  Lav, 

3 

57 

42 

24 

189 

30 

2M 

8th 

0  W,  C. 
OU. 

00 
00 

000 
000 

34 

1650 

10 

165 

6W.  C. 

18 

108 

4  Lav. 

3 

12 

29 

177 

38 

67 

232 

3.3 

4 

Riser  No.  3 


7th  and 

GW.  C. 

50 

300 

40 

120 

below 

5S.  S, 

4 

20 

4  Lav. 

3 

12 

9 

32 

50 

16 

136 

30 

2M 

8th 

0  W.  C. 

00 

000 

40 

120 

6 

300 

1  S,  S. 

4 

4 

10 

36 

48 

17 

137 

3.3 

3 

612 


CHAPTER  35 — WATER  SUPPLY  PIPING 


The  last  method  is  usually  the  most  satisfactory.  (See  Fig.  6.) 
In  the  first  instance  the  up-feed  riser  may  be  sized  for  the  same  pressure 
drop  as  used  for  the  cold  water  riser  and,  from  the  top  of  the  riser  just 
below  the  top  fixture  connection,  a  return  circulation  line  may  be  carried 
back  to  the  main  return  line  in  the  basement  and  connected  through  a 
check  valve,  set  on  a  45-deg  angle,  and  a  gate  valve;  these  return  circu- 
lation lines  should  never  be  less  than  %  in.,  and  on  the  farther  half  of  the 
risers,  not  less  than  1  in.  to  favor  circulation  in  the  far  end.  Typical  top 
and  bottom  connections  for  such  risers  are  shown  in  Fig.  7. 


4  W.  C.-F.V. 
2U.-F.V. 


8th. 

i*-*- 

7th. 

,      4  W.  C.-F.V. 
24      2U.-F.V. 
J^     3  Lav. 

6th,        : 

4  W.  C.-F.V. 
2j       2U.-F.  V. 
,t          3  Lav. 

5th.        : 

„,     4W.C.-F.V. 
24      2U.-F.V. 

*-*.     3LaV< 

3"       10  Lav. 

4th.        : 

H*~ 

3rd.        : 

3"       1  S.  S. 

2nd.       : 

3"       1  S.  S. 

3"       1  S.  S, 

1st         : 

*-»- 

3"    3"  Main 

(I) 

FIG.  5.    UP-FEED 
SYSTEM 

For  the  second  arrangement  of  hot  water  risers,  circulation  lines  are 
run  back  from  the  last  fixture  supplied  to  the  main  return  circulation  line 
in  the  same  manner  as  just  described,  using  %  in.  for  the  near  risers  and 
1  in.  for  the  far  risers.  The  sizing  is  much  more  difficult,  as  it  is  necessary 
to  start  at  the  bottom  floor  of  the  return  riser  and  work  back  to  the  top  of 
this  riser  and  then  carry  the  maximum  flow  across  onto  the  top  of  the 
corresponding  supply  riser  and  work  down  on  this  riser  from  the  top  floor 
to  the  bottom.  Naturally  this  gives  a  much  greater  flow  in  the  supply 
riser  and  aids  circulation  by  reducing  pipe  friction.  The  allowable  loss 
per  100  ft  in  such  lines  must  be  made  about  half  that  used  for  the  cold 

613 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


Vent 


I! 


// 


-f 

\ 

•*• 

1 

""*" 

| 

-** 

I 

T-.jfr_ 

— 

II 

/•?-• 


(a) 


(6) 


(c) 


FIG.  6.    METHODS  OF  ARRANGING  HOT  WATER  CIRCULATION  LINES 


water  risers  which  do  not  have  the  combined  up-  and  down-travel  which 
the  hot  water  must  make. 

In  the  third  and  most  common  arrangement  all  of  the  water  is  carried 
from  the  tank  or  heater  directly  to  the  top  of  the  building  and  is  there 
distributed  to  the  risers  which  are  down-feed  and  may  be  sized  in  the 
regular  down-feed  manner  if  the  total  equivalent  run  either  from  the 
street  main  or  house  tank  is  taken  into  consideration.  The  return 
circulation  lines  frorn  the  botton  of  each  riser  should  be  arranged  in  the 
manner  already  outlined  and  any  riser  not  going  to  the  basement  to 

TABLE  9.   SUMMARY  OF  RISER  SIZES  TO  GIVEN  MAIN  SIZES  (SEE  FIG.  4) 


RISER 
No. 

FIXTURES 

GPM  PER 

FIXTURE 

RISER 

GPM 

PROBABLE 
USE 
(PER  CENT) 

PROBABLE 

GPM 

ALLOWABLE 
DROP  (La 
PER  100  FT) 

SIZE  OF 

MAIN 
(INCHES) 

1 

12  W.  C. 
6U. 

50 
40 

600 
240 

18 
37 

151 
69 

18 
31  Ftxt. 

840 
186 

220 

3.3 

4 

2 

29  W.  C. 
5U. 

50 
40 

1450 
200 

8 

199 

52 
29  Fixt. 

2490 
177 

00 

363 

33 

120 

319 

3.3' 

4 

3 

6  W,  C. 

50 

300 

7 
33 

195 
131 

58 
10  Fixt. 

2790 
36 

70 

399 

326 

3.3 

4 

614 


CHAPTER  35 — WATER  SUPPLY  PIPING 


TABLE  10.    TYPICAL  CALCULATION  OF  PIPE  SIZES  ON  UP-FEED  RISER  WITH 
FLUSH  VALVE  WATER-CLOSETS  AND  URINALS  (SEE  FIG.  5) 


FLOOR 

FIXTURES 

GPM 

MAXIMUM 

PROBABLE 

PROBABLE 

PROBABLE 

ALLOWABLE 

PIPE 

OF 

ON 

PER 

FIXTURES 

USE 

FIXTURES 

RISER 

DROP 

SIZE 

BLDG. 

FLOOR 

FIXTURE 

GPM 

(PER  CENT) 

GPM 

GPM 

LB  PER  100  FT 

IN. 

Riser  No.  1 


8th 

4  W.  C. 
2U. 

6 
3  Lav. 

50 
40 

•   3 

200 
80 

~280 
9 

40 
80 

112 
7 

119 

8.5 

zy* 

7th 

4  W.  C. 

2U. 

50 
40 

200 

80 

12 
3  Lav. 

6 

3 

560 
9 

_ 

25 
55 

140 
10 

150 

8.5 

2H 

6th 

4  W.  C. 

2U. 

50 

40 

200 

80 

18 
3  Lav. 

3 

840 
9 

18 

151 

0 

27 

50 

14 

165 

8.5 

1\i 

5th 

4  W.  C. 
2U. 

50 

40 

200 
80 

24 
3  Lav. 

3 

1120 
9 

15 

168 

12 

36 

47 

17 

185 

8.5 

3 

4th 

24  W.  C." 
andU. 
10  Lav. 

3 

1120 
30 

15 

108 

22 

06 

40 

27 

195 

8.5 

3 

3rd 

24  W.  O 
andU. 
1  S,  S, 

4 

1120 
4 

15 

168 

23 

70 

40 

28 

196 

8,5 

3 

2nd 

24  W.  C." 

andU. 
IS,  S. 

24   ~~ 

4 

1120 

4 

74 

15 

40 

168 
30 

198 

8.5 

3 

1st 

24  W.  C.« 

andU. 

is.  a 

4  ' 

1120 

4 

15 

168 

25 

78 

40 

31 

199 

8.5 

3 

«Frora  floors  above. 


615 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


supply  fixtures  must  have  these  returns  carried  down  to  the  basement 
from  the  termination  of  the  supply  riser  at  whatever  level  it  may  end. 
All  risers,  both  hot  and  cold,  should  be  valved  at  the  main  with  an 
extra  check  valve  on  the  hot  water  return  circulation  so  that  the  risers 
may  be  cut  off  and  repaired  when  necessary  without  disturbing  the 
service  in  the  remainder  of  the  system. 

HOT  WATER  SUPPLY 

Having  designed  the  service  hot  water  piping,  the  next  step  is  to  furnish 

Top  Fixture  Connection  and  Air  Vent-% 

'' 


t 


'S* 


'•Return  Main         '-Supply  Main 

FIG.  7.    SUPPLY  AND  RETURN 

MAIN  CONNECTIONS  FOR  HOT 

WATER  SUPPLY  SYSTEM 

some  means  of  heating  the  water  and  in  this  respect  it  is  necessary  to  pass 
from  the  maximum  probable  flow  to  the  maximum  probable  hourly 
demand,  which  is  quite  different.  If  an  instantaneous  heater  were  used, 
it  would  require  adequate  capacity  to  provide  for  the  heating  of  the  water 
as  fast  as  it  is  drawn  and  a  heater  of  this  type  should  be  sized  on  the  basis 
of  the  maximum  probable  flow  with  the  accompanying  heavy  drafts  on 
the  heating  device  and  with  intervals  of  no  draft  at  all.  To  balance  these 
inequalities  of  flow  the  storage-type  heater  is  often  utilized  so  that  the 
water  demand  can  be  heated  during  periods  of  light  demand  and  stored 
up  for  use  during  the  periods  of  heavy  demand.  The  total  water  con- 
sumption per  person  usually  varies  between  100  and  150  gal  per  day  when 

616 


CHAPTER  35 — WATER  SUPPLY  PIPING 


laundry  and  culinary  operations  for  the  occupants  are  carried  out  on  the 
same  premises.  The  maximum  hourly  demand  under  these  conditions 
will  be  found  to  be  about  one-tenth  of  the  average  daily  consumption. 
If  one-third  of  the  total  water  used  is  hot  water  and  125  gal  per  day  is 
assumed  as  a  fair  average  of  consumption  per  person,  it  is  apparent  that 
each  person  uses  about  40  gal  of  hot  water  per  day.  If  one-  tenth  of  this 
represents  the  peak  hourly  load,  then  4  gph  must  be  allowed  per  person 
for  the  heaviest  demand.  If  the  average  occupancy  of  apartments  is 
3  persons,  the  peak  hour  demand  per  apartment  will  be  about  12  gph.  It 
is  customary  to  allow  10  gph  of  heating  capacity  per  apartment.  Water 
in  excess  of  this  heating  capacity  drawn  out  during  the  peak  hours  is 
provided  for  by  storage  in  the  hot  water  tank  where  this  water  is  heated 
during  hours  when  the  demand  is  below  the  average. 

HOT  WATER  STORAGE 

The  amount  of  storage  provided  in  the  hot  water  tank  or  heater  is 
somewhat  a  matter  of  choice  but  is  usually  made  ample  to  carry  over  the 
peak  shortage  which  is  likely  to  occur  and  is  based  on  the  assumption  that 
only  75  per  cent  of  the  storage  capacity  will  be  available,  as  it  has  been 
found  that  if  more  than  this  amount  is  withdrawn  from  storage,  the  tank 
is  so  cooled  down  as  to  make  the  balance  useless.  The  general  rule  may 
be  cited  that  the  less  the  heating  capacity  the  greater  must  be  the  storage, 
and  the  greater  the  storage  the  less  may  be  the  heating  capacity  down  to  a 
point  where  the  heating  capacity  will  fail  to  be  sufficient  to  heat  up  the 
tank  storage  during  the  periods  of  small  load. 

Example  5.    A  heater  to  supply  500  persons  will  have  an  average  daily  use  of  about 

500  X  40  gal  =  20,000  gal 
and  this  is  an  average  of 


but  the  peak  hour  will  require 

Ho  of  20,000  =  2,000  gal 

and  the  shortage  during  the  peak  hour,  if  the  heating  capacity  is  made  to  suit  the  average 
hourly  use  of  833  gal,  will  be 

2,000  -  833  -  1167  gal 

so  that  the  storage  capacity,  based  on  75  per  cent  being  available  from  this  capacity 
without  cooling  the  tank  excessively,  will  be 

1167       1KKr     . 
-  1556  gal 


Should  it  be  desired  to  reduce  the  size  of  storage  tanks  and  to  use  a  greater  heating 
capacity,  it  is  only  necessary  to  increase  the  heating  capacity  to  say  1200  gph  which  then 
gives 

2,000  -  1200  «  800  gal 

as  the  shortage  during  the  peak  hour,  and  the  necessary  storage  will  be 

*$.•!?-"""+> 

or  the  heating  capacity  can  be  increased  to  1500  gal,  leaving  a  shortage  of 

2000  -  1500  «  500  gal 
617 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

TABLE  11.    ORDINARY  MAXIMUM  HOURLY  DEMAND  FOR  HOT  WATER  FOR  VARIOUS 
FIXTURES  IN  GALLONS  AND  PROBABLE  PERCENTAGE  OF  USAGE 


TTPE  OF 
BUILDING 

LAVATORIES 

BATHS 

SHOWEES 

SLOP 
SINKS 

KITCHEN 

SINKS 

PANTRY 

SINKS 

FOOT 
BATHS 

WASH 

TRAtS 

Av. 

MAX 
USE* 

Private 

Public 

MAXIMUM 
PROBABLE 
USAGE 

GPM 

20 

20 

40 

300 

30 

30 

20 

20 

50 

Probable  Usage  in  Per  Cent  of  Maximum  Ordinary  Use 

35 

60 
80 
45 
70 
90 
100 
20 
100 
50 
25 
75 

Apt.  house 
Club 
Gym. 
Hospital 
Hotel 
Industrial 
Laundries 
Office  building 
Baths 
Residences 
Schools 
Y.  M.  C.  A. 

25 
25 
25 
25 
25 
25 
25 
25 
25 
25 
25 
25 

50 
75 
100 

75 
100 
150 
100 
75 
150 

"75 
100 

33 
50 
100 
50 
50 
100 

150 
50 

166 

67 
67 
100 
33 
33 
100 

100 
33 
100 
100 

67 
67 

"67 
100 
67 
33 
50 
50 
50 
67 
67 

33 

67 

67 
67 
67 

50 
100 

100 
100 

25 
25 
100 
25 
25 
100 

60 
80 

"so 

80 
100 

33 
33 

67 

50 
100 
100 

50 
50 
100 

60 

"so 

aPercentage  of  fixtures  likely  to  be  demanding  maximum  probable  usage  at  any  one  time. 

TABLE  12.    HOT  WATER  CONSUMPTION  IN  VARIOUS  TYPES  OF  BUILDINGS 
FOR  DIFFERENT  PURPOSES 


TYPB  OF  BOTLDINO 

CONDITIONS 

GALLONS 

Hotels 

Room  with  basin  only 
Room  with  bath 
(Transient) 
(Men) 
(Mixed) 
(Women) 
Two-room  suite  and  bath 
Three-room  suite  and  bath 

10  (per  day) 

40  (per  day) 
40  (per  day; 
60  (per  day 
80  (per  day 
80  (per  day 
100  (per  day 

Public 
Buildings 

Public  bath  or  lavatory 
Public  shower 
Public  lavatory  with  attendant 

150  (per  day  per  fixture)  . 
200  (per  day  per  fixture) 
200  (per  day  per  fixture) 

Industrial 
Buildings 

Per  office  employee 
Per  factory  employee 
Cleaning  floors 

2  (per  day) 
5  (per  day) 
3  (per  1000  sq  ft  per  day) 

Restaurants 

S0.50  Meals 
$1.00  Meals 
$1.50  Meals 

0.5  (per  customer  with  hand  washing) 
1.0  (per     customer     with     machine 
washing) 
1,0  (per  customer  with  hand  washing) 
2.0  (per     customer     with     machine 
washing) 
1.5  (per  customer  with  hand  washing) 
4.0  (per    customer    with     machine 
washing) 

618 


CHAPTER  35 — WATER  SUPPLY  PIPING 


and  the  storage  required  only 


Good  design  requires  that  the  heating  capacity  be  made  as  small  as 
possible  without  introducing  undesirable  amounts  of  storage,  as  the 
heating  capacity  directly  determines  the  load  on  the  source  of  heat. 

As  indicated  in  Example  5,  the  heating  load  is  proportional  to  the 
heating  capacity  and  the  boiler  capacity  must  be  increased  for  higher 
heating  capacities  and  may  be  reduced  for  smaller  heating  capacities  with 
greater  storage.  It  may  be  assumed  that  a  boiler  capacity  of  about  3J^ 
sq  ft3  of  equivalent  steam  heating  surface  (radiation)  must  be  provided 
for  every  gallon  of  water  heated  100  F  or  from  50  F  to  150  F,  which  is 
the  temperature  rise  most  commonly  assumed  and  required.  On  this 
basis  it  will  be  seen  that  the  various  conditions  cited  in  Example  5  will 
require  additional  boiler  capacity  as  follows: 

Heating  Capacity  Additional  Boiler  Capacity 
(gph)  (Sq  Ft  EDR) 

833  2916 

1200  4200 

1500  5250 

From  this  it  is  apparent  that  it  is  less  costly  to  provide  ample  storage 
and  to  reduce  boiler  capacity  than  to  diminish  the  storage  and  supply  a 
greatly  increased  boiler  capacity  to  compensate. 

ESTIMATING  HOT  WATER  DEMAND  BY  FIXTURES 

In  buildings  where  the  occupancy  is  doubtful  and  only  the  number  of 
plumbing  fixtures  can  serve  as  a  basis  for  determining  the  probable  hot 
water  demand,  the  problem  is  not  so  simple  owing  to  the  fact  that  a 
fixture  gives  no  information  as  to  how  heavy  a  service  may  be  demanded 
from  the  fixture  and  this  amount  of  service  is  really  the  governing  factor 
in  making  an  estimate  of  the  probable  hot  water  demand.  Table  11  may 
prove  of  some  value  in  this  respect  as  it  gives  the  maximum  assumed 
quantity  of  hot  water  per  hour  which  will  be  demanded  of  any  fixture  and 
then  gives  a  percentage  of  this  amount  which  may  be  assumed  as  probable 
in  different  types  of  buildings.  Table  12  gives  approximate  hot  water  re- 
quirements in  various  types  of  buildings. 

Example  6.  Let  it  be  assumed  that  an  apartment  house  with  20  apartments  has  20 
baths,  20  lavatories,  20  kitchen  sinks  and  20  laundry  trays;  what  is  the  probable  maxi- 
mum hourly  demand  for  hot  water? 

20  Baths  at  40  gal  and  33  per  cent  .....  „  .......  .  ........................  .  .......................  270  gal 

20  Lavs,  at  20  gal  and  25  per  cent  ..........  .  ................................................  .,„  100  gal 

20  Sinks  at  30  gal  and  33  per  cent  ....................................  .  ......  .  ..................  200  gal 

20  Trays  at  50  gal  and  60  per  cent  ..............  .  ...............................................  600  gal 


Total 1170  gal 

Probable  peak  use  at  one  time .„ 35  per  cent 

Probable  actual  peak  demand 409  gph 

100  "^   ^  3*^ 
'Actual  requirement  for  100-deg  tempeiature  difference  —  — "ZAQ     '    ""   ***^  9C1  ^  per  Ballon  of 

water  heated. 

519 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

If  three  persons  are  assumed  to  an  apartment  the  total  daily  use  of  hot  water  should 
approximate 

20  X  3  X  40  gal  =  2400  gal 

and  if  the  peak  hour  is  10  per  cent  of  this  amount,  the  peak  hour  by  this  method  shows  a 
probable  demand  of  one-tenth  of  2400  gal,  which  indicates  that  the  values  in  Table  7 
are  safe. 

PROBLEMS  IN  PRACTICE 

1  •  Why  is  it  impractical  to  size  water  supply  piping  so  pipe  friction  will  pro- 
duce an  equal  pressure  on  each  fixture? 

Because  the  friction  would  be  built  up  only  in  periods  of  maximum  flow  and  at  all  other 
times  it  would  be  only  a  fraction  of  that  required. 

2  •  What  is  the  purpose  of  zoning  water  supply  systems  in  tall  buildings? 

To  avoid  excessive  pressures  in  the  lower  stories. 

3  •  Define  the  maximum  possible  flow,  the  maximum  probable  flow,  and  the 
average  probable  flow. 

The  maximum  possible  flow  is  the  flow  which  would  occur  if  all  of  the  outlets  on  the 
system  were  opened  at  one  and  the  same  time.  The  maximum  probable  flow  is  the  flow 
which  will  occur  with  probable  peak  conditions.  The  average  probable  flow  is  the  flow 
likely  to  occur  under  a  normal  condition  of  use. 

4  •  What  is  the  factor  of  usage? 

This  is  the  percentage  of  the  maximum  possible  flow  which  is  likely  to  occur  at  peak  load. 

5  •  Within  what  limits  does  the  factor  of  usage  lie? 

From  100  per  cent  for  a  single  fixture  down  to  about  28  per  cent  for  1000  fixtures  of 
ordinary  type,  and  from  100  per  cent  for  a  single  fixture  down  to  about  1  per  cent  for 
1000  fixtures  of  the  flush  valve  type. 

6  •  How  many  feet  higher  than  the  uppermost  fixtures  must  the  water  line  in 
a  house  tank  be  to  provide  about  15  Ib  per  square  inch  pressure  at  the  fixture 
outlet? 

Allowing  for  pipe  losses,  about  45  ft. 

7  •  What  methods  of  hot  water  circulation  commonly  are  employed  with  hot 
water  supply  systems? 

a.  Upfeed  risers  with  returns  having  no  connections  paralleling  the  risers. 

b.  Upfeed  risers  with  returns  in  other  locations,  and  with  connections  taken  off  both 
supply  and  return. 

c.  One  main  upfeed  riser,  without  connections,  supplying  all  downfeed  risers  for  all 
fixtures. 

ft  •  Which,  method  of  hot  water  supply  generally  is  the  most  satisfactory? 

The  single  main  upfeed  riser  supplying  drop  risers  for  all  fixtures. 

9  •  How  much  of  the  water  stored  in  a  hot  water  storage  tank  really  is  available 
for  use? 

About  75  per  cent,  because  when  only  25  per  cent  of  the  original  water  remains  in  the 
tank  it  has  been  so  cooled  clown  by  the  entering  water  that  it  is  too  cold  for  satisfactory 
use. 

10  •  In  cases  of  intermittent  demand,  does  a  large  hot  water  storage  tank 
increase  or  decrease  the  steam  load  for  water  heating? 

It  decreases  the  steam  load  in  cases  of  intermittent  demand  but  causes  no  change  in  the 
steam  load  if  the  demand  is  constant, 

620 


Chapter  36 

INSULATION  OF  PIPING 

Heat  Losses  from  Bare  Pipes,  Steam  and  Hot  Water  Lines,  Low  Tem- 
perature   Pipe    Insulation,   Pipe    Sweating,    Heat    Losses  from    Pipe 
Surfaces,  Thickness  of  Pipe  Insulation,  Underground  Insulation 

PIPE  insulation  performs  an  important  function  in  preventing  loss  of 
heat  where  steam  or  hot  water  are  conveyed  from  one  part  of  a 
building  to  another,  and  in  reducing  the  absorption  of  heat  by  cold  pipes 
as  well  as  preventing  condensation  on  the  outer  surfaces. 

BARE  PIPE  LOSSES 

Heat  losses  from  horizontal  bare  iron  pipes,  based  on  data  obtained 
from  tests  conducted  at  the  Mellon  Institute,  are  given  in  Table  1.  These 
losses  are  expressed  in  Btu  per  hour  per  linear  foot  of  pipe  per  degree 
Fahrenheit  difference  in  temperature  between  the  steam  or  hot  water  in 
the  pipe  and  the  air  surrounding  the  pipe.  The  monetary  value  of  the 
loss  of  heat  given  in  Table  1  may  be  obtained  by  means  of  Fig.  1  for 
various  heating  system  efficiencies,  temperature  differences,  and  calorific 
values  and  costs  of  coal.  To  solve  a  problem,  select  the  proper  heat  loss 
coefficient  from  Table  1  and  locate  this  value  on  the  upper  left  hand 
margin  of  the  chart.  Then  draw  lines  in  the  order  indicated  by  the 
dotted  lines,  the  dollar  value  of  the  heat  loss  per  100  linear  feet  of  pipe  per 
1000  hours  being  given  on  the  upper  right  hand  scale.  In  using  this 
chart,  the  cost  of  coal  should  also  include  the  labor  for  handling  it,  boiler 
room  expense,  etc.  For  additional  information  on  this  subject  refer  to 
paper  entitled  Heat  Emission  from  Iron  and  Copper  Pipe1,  by  F.  C. 
Houghten  and  Carl  Gutberlet. 

In  order  to  determine  heat  losses  per  linear  foot  of  pipe  from  known 
losses  per  square  foot,  it  is  necessary  to  know  the  area  in  square  feet  per 
linear  foot  of  pipe.  Table  2  gives  these  areas  for  various  standard  pipe 
sizes  while  Table  3  gives  the  area  in  square  feet  for  flanges  and  fittings 
for  various  standard  pipe  sizes. 

Very  often,  even  where  pipes  are  thoroughly  insulated,  flanges  and 
fittings  are  left  bare  due  to  the  belief  that  the  losses  from  these  parts  are 
not  large.  However,  the  fact  that  a  pair  of  9-in.  standard  flanges  having 
an  area  of  3.00  sq  ft  would  lose,  at  100  Ib  steam  pressure,  an  amount  oi 

lA.S.H.V.B,  TRANSACTIONS,  Vol.  38,  1932. 

021 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


TABLE  1.    HEAT  LOSSES  FROM  HORIZONTAL  BARE  IRON  PIPES 

Expressed  in  Btu  per  linear  foot  per  degree  Fahrenheit  difference  in  temperature  between  the 
pipe  and  surrounding  still  air  at  70  F 


HOT  WATER 

STEAM 

NOMINAL 
PGPB 

120  F 

150  F 

180  F 

210  F 

227.1  F 
(SLb) 

297.7  F 
(50  Lb) 

337.9  F 
(100  Lb) 

SIZE 

(INCHES) 

TEMPERATURE  DIFFERENCE 

50  F 

80  F 

110  F 

140  F 

157.1  F 

227.7  F 

267.9  F 

Ji 

0.543 

0.573 

0.605 

0.638 

0.656 

0.742 

0.796 

% 

0.660 

0.690 

0.729 

0.762 

0.781 

0.886 

0.955 

1 

0.791 

0.829 

0.878 

0.920 

0.953 

1.084 

1.166 

1J4 

0.979 

1.02 

1.087 

1.15 

1.184 

1.345 

1.450 

1J/| 

1.09 

1.15 

1.220 

1.29 

1.335 

1.520 

1.640 

2 

1.34 

1.40 

1.491 

1.58 

1.637 

1.866 

2.015 

2y2 

1.58 

1.67 

1.778 

1.87 

1.937 

2.215 

2.388 

1.88 

1.99 

2.100 

2.22 

2.301 

2.641 

2.853 

v/* 

2.13 

2.24 

2.3SO 

2.51 

2.585 

2.972 

3.215 

4: 

2.36 

2.50 

2.650 

2.78 

2.873 

3.312 

3.582 

4M 

2.60 

2.75 

2.920 

3.08 

3.170 

3.655 

3.956 

2.87 

3.02 

3.200 

3.38 

3.493 

4.030 

4.368 

6 

3.39 

3.56 

3.775 

4.01 

4.115 

4.755 

5.153 

8 

4.32 

4.55 

5.050 

5.14 

5.270 

6.120 

6.635 

10 

5.32 

5.61 

5.925 

6.34 

6.551 

7.592 

8.245 

12 

6.25 

6.62 

6.995 

7.46 

7.670 

8.900 

9.670 

TABLE  2.    RADIATING  SURFACE  PER  LINEAR  FOOT  OF  PIPE 


NOMINAL 

SURFACE 

NOMINAL 

SURFACE 

NOMINAL 

SURFACE 

PIPE  SIZE 

AREA 

PIPE  SIZB 

AREA 

PIPE  SIZE 

AREA 

(INCHES) 

(SqFr) 

(INCHES) 

(So  FT) 

(INCHES) 

(So  FT) 

y% 

0.22 

2 

0.622 

5 

1.456 

% 

0.275 

2H 

0.753 

6 

1.734 

i 

0.344 

3 

0.917 

8 

2.257 

iji 

0.435 

3M 

1.047 

10 

2.817 

i« 

0.498 

1.178 

12 

3.338 

TABLE  3.    AREAS  OF  FLANGED  FITTINGS,  SQUARE  FEET* 


NOMINAL 

FLANGED 
COUPLING 

90  DEQ  ELL 

LONO  RADIUS 
ELL 

TEE 

CROSS 

PIPE  SIZE 

(INCHES) 

Standard 

Extra 
Heavy 

Standard 

Extra 
Heavy 

Standard 

Extra 
Heavy 

Standard 

Extra 
Heavy 

Standard 

Extra 
Heavy 

1 

0,320 

0,438 

0.795 

1.015 

0.892 

1.083 

1.235 

1.575 

1,622 

2,07 

1« 

0,383 

0.510 

0.957 

1.098 

1.084 

1.340 

1.481 

1,925 

1.943 

2.53 

1« 

0.477 

0,727 

1,174 

1.332 

1.337 

1.874 

1,815 

2.68 

2.38 

3.54 

2 

0.672 

0.848 

1,65 

2.01 

1.84 

2,16 

2,54 

3,09 

3.32 

4,06 

2^ 

0.841 

1.107 

2.09 

2.57 

2.32 

2,76 

3.21 

4.05 

4.19 

5.17 

3 

0.945 

1.484 

2.38 

3.49 

2.68 

3.74 

3.66 

5.33 

4.77 

6.95 

3^2 

1.122 

1.644 

2.98 

3.96 

3.28 

4,28 

4,48 

6.04 

5.83 

7.89 

4 

1.344 

1.914 

3.53 

4.64 

3,96 

4.99 

5,41 

7.07 

7,03 

9,24 

Q4 

1.474 

2.04 

3.95 

5,02 

4.43 

5.46 

6.07 

7.72 

7.87 

10.07 

1.622 

2.18 

4.44 

5.47 

5.00 

6.02 

6.81 

8.52 

8.82 

10.97 

6 

1.82 

2.78 

5,13 

6.99 

5.99 

7.76 

7.84 

10.64 

10,08 

13.75 

8 

2.41 

3.77 

6,98 

9.76 

8.56 

11.09 

10.55 

14.74 

13,44 

18.97 

10 

3,43 

5,20 

10.18 

13.58 

12,35 

15.60 

15,41 

20.41 

19.58 

26.26 

12 

4,41 

6.71 

13,08 

17.73 

16.35 

18,76 

19,67 

26,65 

2487 

34,11 

•Including  areaaol  accompanying  flanges  bolted  to  the  fitting, 

622 


CHAPTER  36 — INSULATION  OF  PIPING 


FIG,  1.    CHART  FOR  ESTIMATING  DOLLAR  VALUE  OF  HEAT  Loss 
FROM  BARE  IRON  PIPES.   (SEE  TABLE  1)* 

aThls  chart  is  based  on  100  linear  feet  per  1000  hours.  For  fractions  or  multiples  of  these  factors, 
multiply  by  proper  percentage, 

heat  equivalent  to  more  than  a  ton  of  coal  per  year  shows  the  necessity 
for  insulating  such  surfaces.  Table  3  shows  the  areas  of  both  standard 
and  extra  heavy  flanged  fittings  including  the  accompanying  flanges 
bolted  to  the  fittings, 

STEAM  AND  HOT  WATER  LINES 

The  conductivities  of  various  materials  used  for  insulating  steam  and 
hot  water  pipes  are  given  in  Table  4.  In  this  table  the  conductivities  are 
given  as  functions  of  the  mean  temperatures  or  the  mean  of  the  inner  and 

623 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


TABLE  4,    CONDUCTIVITIES  (jfe)  OF  VARIOUS  TYPES  OF  INSULATING  MATERIALS 
FOR  MEDIUM  AND  HIGH  TEMPERATURE  PipESa 

MBAN  TEMPERATURE 


100  F 

200  F 

300  F 

400  F 

500  F 

85  per  cent  Magnesia,  Type 

0.425 

0.465 

0.505 

0.550 

0.590 

Corrugated  Asbestos  Type 

0.530 

0.650 

0.770 

0.890 

.(4  Plies  per  1  in.  thick) 
Corrugated  Asbestos  Type 

0.480 

0.555 

0.630 

0.705 

(8  Plies  per  1  in.  thick) 
Laminated  Asbestos  Type        ...       .  . 

0.360 

0.415 

0.470 

0.525 

0.585 

(30-40  Laminations  per  1  in.  thick) 
Laminated  Asbestos  Type        

0.545 

0.605 

0.665 

0.725 

0.785 

(20  Laminations  per  1  in.  thick) 
Rock  Wool  Type      .                ..... 

0.350 

0.410 

0.470 

0.530 

0.590 

High  Temperature  Type 

0.515 

0.545 

0.575 

0.605 

0.635 

(Diatomaceous  Earth  and  Asbestos) 
Brown  Asbestos  Type 

0.600 

0.640 

0.675 

0.715 

0.750 

(Felted  Fibre) 

^Mechanical  Engineers'  Handbook,  Marks,  3rd  Ed.,  1930. 

outer  surface  temperatures  of  the  insulations.  This  method  of  stating 
conductivities  makes  it  possible  readily  to  calculate  the  heat  loss  through 
single  or  compound  sections.  It  should  be  emphasized  that  the  con- 
ductivities given  in  Table  4  for  the  various  insulations  are  the  average  of 
values  obtained  from  a  number  of  tests  made  on  each  type  of  material, 
also  that  all  variables  due  to  differences  in  thickness,  pipe  sizes,  and  air 
conditions  are  eliminated.  Individual  manufacturer's  materials  will,  of 
course,  vary  in  conductivity  to  some  extent  from  these  values. 

The  heat  losses  through  six  of  the  types  of  insulation  given  in  Table  4 
for  1,  1J^  and  2-in. -thick  materials,  and  for  temperatures  commonly 
encountered  in  engineering  practice  can  be  obtained  from  Tables  5  to  10, 
inclusive.  The  loss  through  other  thicknesses  of  the  materials,  and  for 
other  hot  water  or  steam  temperature  conditions  may  be  obtained  by 
interpolation.  The  heat  loss  coefficients  given  in  Tables  5  to  10  are 
based  on  the  conductivities  in  Table  4  and  were  computed  from  data 
given  in  Chapter  22,  THE  GUIDE  1931. 

LOW  TEMPERATURE  PIPE  INSULATION 

Surfaces  maintained  at  low  temperatures  should  be  insulated  so  as  to 
retard  the  flow  of  heat  from  the  outside  into  the  low  temperature  area  and 
to  prevent  the  formation  of  condensation  and  of  frost  if  the  temperatures 
are  low  enough,  as  well  as  to  prevent  corrosion  induced  by  the  presence 
of  condensed  moisture  on  metal  surfaces.  Materials  commonly  used  for 
insulating  pipes  and  surfaces  at  low  temperatures  are  cork,  rock  cork, 
hair  felt  and  other  felted  or  fibrous  non-absorbent  materials.  Thermal 
conductivities  of  low  temperature  insulating  materials  are  given  in 
Chapter  5. 

Insulating  materials  are  available  commercially  to  meet  varying  tem- 
perature gradients.  For  example,  the  thickness  of  insulation  for  ice  water 
is  approximately  1J^  in,  if  the  temperature  in  the  line  is  not  lower  than 

624 


CHAPTER  36 — INSULATION  OF  PIPING 


TABLE  5.    COEFFICIENTS  OF  TRANSMISSION  (U)  FOR  PIPES  INSULATED 
WITH  85  PER  CENT  MAGNESIA  TYPE  INSULATION 

These  coefficients  are  expressed  in  Btu  per  hour  per  square  foot  of  pipe  surface  per  degree 
Fahrenheit  difference  in  temperature  between  pipe  and  surrounding  still  air  at  70  F 


HOT  WATER                                                  STEAM 

THICKNESS 

NOMINAL 

II 

OF 

INSULATION 

PIPE 

SIZE 

120  F 

1SOF 

180  F 

210  F     I    22  VJ 
1     (5  Lb) 

297.7  F 
(50  Lb) 

337.9  F 
(100  Lb) 

(INCHES) 

(INCHES) 

TEMPERATURE  DIFFERENCE 

50  F 

80  F 

110  F 

140  F 

157.1  F 

227.7  F 

267.9  F 

y2 

0.744 

0.754 

0.764 

0.774 

0.779 

0.802 

0.814 

H 

0.672 

0.681 

0.689 

0.697 

0.701 

0.721 

0.731 

i 

0.613 

0.621 

0.629 

0.637 

0.641 

0.659 

0.670 

U4 

0.562 

0.570 

0.577 

0.585 

0.589 

0.606 

0.617 

ix 

0.532 

0.539 

0.546 

0.553 

0.557 

0.573 

0.582 

2 

0.500 

0.506 

0.512 

0.519 

0.523 

0.538 

0.547 

11A 

0.475 

0.481 

0.487 

0.493 

0.497 

0.512 

0.520 

3 

0.455 

0.461 

0.467 

0.474 

0.477 

0.492 

0.500 

1 

3H 

0.441 

0.447 

0.452 

0.458 

0.462 

0.475 

0.483 

4 

0.429 

0.435 

0.441 

0.446 

0.449 

0.463 

0.471 

4^ 

0.420 

0.425 

0.431 

0.437 

0.440 

0.453 

0.460 

5 

0.411 

0.416 

0.422 

0.427 

0.430 

0.443 

0.450 

6 

0.402 

0.408 

0.413 

0.419 

0.422 

0.435 

0.442 

8 

0.387 

0.392 

0.397 

0.403 

0.405 

0.418 

0.425 

10 

0.375 

0.380 

0.385 

0.390 

0.393 

0.405 

0.412 

12 

0.369 

0.374 

0.378 

0.383 

0.386 

0.398 

0.405 

1A 

0.617 

0.625 

0.633 

0.642 

0.646 

0.665 

0.676 

H 

0.550 

0.558 

0.566 

0.573 

0.577 

0.596 

0.606 

i 

0.4% 

0.503 

0.511 

0.518 

0.522 

0.540 

0.549 

ilA 

0.453 

0.459 

0.465 

0.472 

0.475 

0.490 

0.498 

IX 

0.424 

0.430 

0.436 

0.442 

0.445 

0.459 

0.467 

2 

0.394 

0.400 

0.405 

0.410 

0.413 

0.427 

0.434 

11A 

0.371 

0.376 

0.382 

0.386 

0.389 

0.401 

0.408 

3 

0.352 

0.357 

0.362 

0.367 

0.370 

0.380 

0.387 

IX 

&A 

0.339 

0.343 

0.347 

0.351 

0.354 

0.364 

0.370 

4 

0.328 

0.333 

0.337 

0.341 

0.343 

0.353 

0.359 

4M 

0.320 

0.324 

0.328 

0.332 

0,334 

0.343 

0.350 

5 

0.312 

0.316 

0.320 

0.324 

0.326 

0.336 

0.342 

6 

0.303 

0.307 

0.311 

0.315 

0.318 

0.328 

0.333 

8 

0.287 

0.291 

0.295 

0.299 

0.301 

0.311 

0.316 

10 

0.276 

0.280 

0.284 

0.288 

0.290 

0.299 

0.304 

12 

0.272 

0.275 

0.279 

0.283 

0,285 

0.294 

0.299 

X 

0.543 

0.551 

0.558 

0,565 

0.569 

0.587 

0.597 

K 

0.484 

0.490 

0.497 

0.503 

0.507 

0.523 

0.532 

0.433 

0.439 

0.445 

0.451 

0.454 

0.467 

0.476 

11A 

0.393 

0.398 

0.403 

0.409 

0.412 

0.424 

0.432 

IX 

0.365 

0.370 

0.376 

0.381 

0.384 

0.397 

0.402 

2 

0,338 

0.343 

0.347 

0.351 

0.354 

0.364 

0.370 

2J^ 

0.316 

0.320 

0,324 

0.328 

0.331 

0.341 

0.347 

3 

0.297 

0.301 

0.305 

0.309 

0.312 

0.321 

0.326 

2 

3H 

0.284 

0.288 

0.292 

0.295 

0.297 

0.306 

0.311 

4 

0.275 

0.278 

0.282 

0.285 

0.287 

0.296 

0.301 

42^ 

0.266 

0.270 

0.273 

0.276 

0.278 

0.286 

0.290 

0.258 

0.262 

0.265 

0.268 

0.270 

0.278 

0.283 

6 

0.250 

0.254 

0.257 

0.260 

0.262 

0.270 

0.274 

8 

0.236 

0.239 

0,242 

0,245 

0.247 

0.255 

0.258 

10 

0.224 

0.227 

0,230 

0.233 

0.235 

0.242 

0.246 

12 

0,219 

0,222 

0.225 

0.228 

0.230 

0,237 

0,240 

625 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


TABLE  6.   COEFFICIENTS  OF  TRANSMISSION  ( U)  FOR  PIPES  INSULATED  WITH  CORRUGATED 
ASBESTOS  TYPE  INSULATION  (4  PLIES  PER  INCH  THICKNESS) 

These  coefficients  are  expressed  in  Btu  per  hour  per  square  foot  of  pipe  surface  per  degree 
Fahrenheit  difference  in  temperature  between  pipe  and  surrounding  still  air  at  70  F 


HOT  WATER 

STEAM 

THICKNESS 

NOMINAL 

OP 

INSULATION 

PIPE 

SIZE 

120  F 

150  F 

180  F 

210  F 

227.1  F 
(5Lb) 

297.7  F 
(50  Lb) 

337.9  F 
(100  Lb) 

(INCHES) 

(INCHES) 

TEMPERATURE  DIFFERENCE 

50  F 

80  F 

110F 

140  F 

157.1  F 

227.7  F 

267.9  F 

1A 

0.890 

0.919 

0.949 

0.978 

0.995 

1.065 

1.106 

% 

0.803 

0.829 

0.857 

0.883 

0.898 

0.961 

0.997 

0.731 

0.756 

0.780 

0.804 

0.818 

0.876 

0.909 

v/± 

0.671 

0.693 

0.716 

0.738 

0.751 

0.804 

0.834 

1H 

0.635 

0.656 

0.677 

0.698 

0.710 

0.760 

0.788 

2 

0.595 

0.615 

0.635 

0.656 

0.667 

0.715 

0.742 

2M 

0.567 

0.586 

0.605 

0.624 

0.635 

0.680 

0.705 

3 

0.544 

0.562 

0.580 

0.598 

0.608 

0.652 

0.677 

1 

3J^ 

0.527 

0.544 

0.561 

0.578 

0.588 

0.631 

0.654 

4 

0.513 

0.530 

0.548 

0.565 

0.575 

0.616 

0.639 

&A 

0.502 

0.518 

0.535 

0.551 

0.561 

0.601 

0.624 

5 

0.490 

0.507 

0.523 

0.539 

0.549 

0.588 

0.611 

6 

0.480 

0.496 

0.512 

0.528 

0.538 

0.577 

0.599 

8 

0.462 

0.477 

0.493 

0.508 

0.517 

0.554 

0.575 

10 

0.447 

0.462 

0.476 

0.491 

0.500 

0.537 

0.557 

12 

0.441 

0.456 

0.470 

0.485 

0.493 

0,529 

0.550 

M 

0.737 

0.762 

0.787 

0.812 

0.826 

0.884 

0.918 

% 

0.657 

0.679 

0.702 

0.725 

0.737 

0.790 

0.820 

0.594 

0.614 

0.634 

0.654 

0.666 

0.713 

0.740 

iy* 

0.542 

0.5s59 

0.577 

0.596 

0.606 

0.649 

0.673 

m 

0.507 

0.524 

0.541 

0.558 

0.568 

0.609 

0.632 

2 

0.471 

0.487 

0.503 

0.519 

0.528 

0.565 

0.587 

% 

0.443 

0.458 

0.473 

0.488 

0.497 

0.533 

0.553 

3 

0.421 

0.435 

0.449 

0.463 

0.472 

0.506 

0.525 

1« 

m 

0.403 

0.417 

0.430 

0.443 

0.451 

0.483 

0.502 

4 

0.393 

0.405 

0.418 

0.432 

0.439 

0.471 

0.489 

4H 

0,383 

0.394 

0.407 

0.420 

0.428 

0.460 

0.476 

5 

0.372 

0.384 

0.397 

0.409 

0.417 

0.447 

0.463 

6 

0.362 

0.374 

0,387 

0.399 

0.406 

0.436 

0.452 

8 

0.343 

0.354 

0.366 

0.378 

0.385 

0.413 

0.429 

10 

0.328 

0.339 

0.351 

0.362 

0.369 

0,397 

0.413 

12 

0.323 

0.334 

0.346 

0.357 

0.364 

0.391 

0.407 

H 

0.648 

0.670 

0.692 

0.713 

0,726 

0.779 

0.810 

H 

0.578 

0,598 

0.617 

0.637 

0.648 

0.694 

0.720 

1 

0.518 

0.535 

0.552 

0.570 

0.580 

0.622 

0.645 

1# 

0.469 

0.485 

0.501 

0.517 

0.527 

0.566 

0.587 

1M 

0.438 

0.452 

0.467 

0.481 

0.490 

0.526 

0.545 

2 

0.404 

0,417 

0.430 

0.444 

0.452 

0.483 

0.502 

2H 

0.379 

0,391 

0.403 

0.415 

0,422 

0.451 

0.466 

3 

0.356 

0.367 

0.378 

0.390 

0.397 

0.425 

0.440 

2 

3^ 

0.339 

0.350 

0.361 

0.373 

0.380 

0.406 

0.421 

4 

0.328 

0.339 

0.350 

0.360 

0.367 

0.392 

0.406 

4M 

0.318 

0.328 

0.339 

0.350 

0.357 

0.381 

0.395 

5 

0.308 

0,318 

0.329 

0.340 

0.346 

0.370 

0.384 

6 

0.299 

0.309 

0.319 

0.329 

0.335 

0.358 

0.371 

8 

0.282 

0.291 

0.301 

0.310 

0.315 

0.336 

0.349 

10 

0,267 

0.276 

0.285 

0.294 

0.299 

0.319 

0.332 

12 

0,263 

0.272 

0,280 

0.289 

0.294 

0,314 

0.325 

626 


CHAPTER  36 — INSULATION  OF  PIPING 


TABLE  7.   COEFFICIENTS  OF  TRANSMISSION  ( U)  FOR  PIPES  INSULATED  WITH  CORRUGATED 
ASBESTOS  TYPE  INSULATION  (8  PLIES  PER  INCH  THICKNESS) 

These  coefficients  are  expressed  in  Btu  per  hour  per  square  foot  of  pipe  surface  per  degree 
Fahrenheit  difference  in  temperature  between  pipe  and  surrounding  still  air  at  70  F 


HOT  WATER 

STEAM 

THICKNESS 

NOMINAL 

OF 

INSULATION 

PIPE 
SIZE 

120  F 

150  F 

180  F 

210  F 

227.1  F 
(SLb) 

297.7  F 
(50  Lb) 

337.9  F 
(100  Lb) 

(INCHES) 

(INCHES) 

TEMPERATURE  DIFFERENCE 

50  F 

80  F 

110  F 

140  F 

157.1  F 

227.7  F 

267.9  F 

}4 

0.801 

0.820 

0.838 

0.857 

0.868 

0.913 

0.939 

%    - 

0.723 

0.739 

0.756 

0.773 

0.783 

0.824 

0.847 

1 

0.658 

0.673 

0.688 

0.704 

0.713 

0.751 

0.772 

1% 

0.606 

0.619 

0.633 

0.647 

0.655 

0.688 

0.707 

m 

0.573 

0.586 

0.599 

0.612 

0.619 

0.652 

0.670 

2 

0.538 

0.550 

0.562 

0.575 

0.581 

0.612 

0.629 

11A 

0.511 

0.523 

0.534 

0.546 

0.553 

0.582 

0.599 

3 

0.489 

0.501 

0.512 

0.524 

0.531 

0.558 

0.575 

1 

3M 

0.474 

0.485 

0.496 

0.507 

0.514 

0.542 

0.557 

4 

0.461 

0.472 

0.482 

0.493 

0.500 

0.527 

0.542 

4M 

0.451 

0.462 

0.472 

0.482 

0.489 

0.515 

0.530 

5 

0.442 

0.452 

0.462 

0.473 

0.479 

0.505 

0.520 

6 

0.432 

0.442 

0.452 

0.463 

0.468 

0.493 

0.508 

8 

0.416 

0.426 

0.436 

0.446 

0.451 

0.475 

0.489 

10 

0.402 

0.412 

0.421 

0.430 

0.435 

0.459 

0.473 

12 

0.397 

0.406 

0.415 

0.424 

0.429 

0.452 

0.466 

1A 

0.664 

0.679 

0.695 

0.711 

0.720 

0.759 

0.780 

% 

0.593 

0.607 

0.621 

0.636 

0.643 

0.677 

0.697 

0.535 

0.547 

0.560 

0.573 

0.580 

0.611 

0,629 

1M 

0.488 

0.499 

0.510 

0.522 

0.528 

0.556 

0.572 

1H 

0.457 

0.467 

0.478 

0.490 

0.496 

0.522 

0.537 

2 

0.425 

0.434 

0.444 

0.455 

0.460 

0.485 

0.499 

m 

0.399 

0.408 

0.418 

0.428 

0.434 

0.457 

0.471 

3 

0.378 

0.387 

0.396 

0.405 

0.411 

0.433 

0.446 

1H 

3^ 

0.363 

0.371 

0.380 

0.388 

0.393 

0.415 

0.427 

4 

0.353 

0.361 

0.369 

0.378 

0.383 

0.403 

0.415 

4^ 

0.343 

0.351 

0.360 

0.368 

0.373 

0.393 

0.404 

5 

0.334 

0.342 

0.350 

0.358 

0.363 

0.383 

0.394 

6 

0.325 

0.333 

0.341 

0.349 

0.353 

0.373 

0.383 

8 

0.309 

0,316 

0.324 

0.332 

0.336 

0.355 

0.365 

10 

0.295 

0.303 

0.310 

0.318 

0.322 

0.340 

0.350 

12 

0.291 

0.298 

0.306 

0.313 

0.317 

0.335 

0,344 

H 

0.585 

0,599 

0.613 

0.627 

0.635 

0.668 

0.688 

H 

0.520 

0.533 

0.545 

0,558 

0.565 

0.595 

0.612 

1 

0.465 

0,476 

0.487 

0.498 

0.504 

0.532 

0.547 

1M 

0.422 

0.432 

0.442 

0.452 

0.458 

0.483 

0.497 

IjLg 

0.394 

0.403 

0.412 

0.422 

0.427 

0.450 

0.462 

2 

0.364 

0.372 

0.380 

0.388 

0.393 

0,415 

0.427 

2H 

0.339 

0.347 

0.355 

0.363 

0.367 

0.387 

0.398 

3 

0.319 

0.327 

0.334 

0.342 

0.346 

0.365 

0.375 

2 

3M 

0.304 

0.311 

0,318 

0.326 

0,330 

0.349 

0.358 

0,295 

0.302 

0.308 

0.315 

0.319 

0.336 

0.345 

4^ 

.0.285 

0.292 

0.299 

0.306 

0.310 

0.327 

0.336 

5 

0.278 

0,284 

0.290 

0.297 

0.301 

0.317 

0.326 

6 

0.269 

0.275 

0,282 

0.288 

0.292 

0.307 

0.315 

8 

0,253 

0.259 

0.265 

0.270 

0.273 

0,288 

0.296 

10 

0.240 

0.245 

0.251 

0,257 

0.260 

0,275 

0,282 

12 

0.236 

0.241 

0.247 

0.2S3 

0.256 

0.270 

0.277 

627 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


TABLE  8.   COEFFICIENTS  OF  TRANSMISSION  (£/)  FOR  PIPES  INSULATED  WITH  LAMINATED 
ASBESTOS  TYPE  INSULATION  (30  TO  40  LAMINATIONS  PER  INCH  THICKNESS) 

These  coefficients  are  expressed  in  Btu  per  hour  per  square  foot  of  pipe  surface  per  degree 
Fahrenheit  difference  in  temperature  between  pipe  and  surrounding  still  air  at  70  F 


HOT  WATER 

STEAM 

THICKNESS 

NOMINAL 

OF 

INSULATION 

PIPE 

SIZE 

120  F 

150  F 

180  F           210  F 

227.1  F 
(SLb) 

297.7  F 
(50  Lb) 

337.9  F 
(100  Lb) 

(INCHES) 

(INCHES) 

TEMPERATURE  DIFFERENCE 

50  F 

80  F 

110  F 

140  F 

157.1  F 

227.7  F 

267.9  F 

H 

0.605 

0.620 

0.635 

0.650 

0.658 

0.695 

0.716 

% 

0.546 

0.560 

0.573 

0.586 

0.594 

0.627 

0.645 

i 

0.498 

0.510 

0.522 

0.534 

0.541 

0.570 

0.587 

M 

0.457 

0.468 

0.480 

0.491 

0.497 

0.525 

0.540 

VA 

0.432 

0.442 

0.453 

0.464 

0.470 

0.496 

0.511 

2 

0.406 

0.416 

0.426 

0.437 

0.442 

0.467 

0.481 

VA 

0.385 

0.395 

0.405 

0.415 

0.420 

0.443 

0.457 

3 

0.370 

0.379 

0.389 

0.398 

0.403 

0.425 

0.438 

1 

m 

0.359 

0.367 

0.376 

0.385 

0.390 

0.413 

0.426 

4 

0.349 

0.358 

0.366 

0.375 

0.380 

0.402 

0.414 

4H 

0.341 

0.350 

0.359 

0.367 

0.372 

0.393 

0.405 

5 

0.334 

0.342 

0.351 

0.359 

0.364 

0.384 

0.395 

6 

0.327 

0.335 

0.343 

0.351 

0.356 

0.376 

0.387 

8 

0.314 

0.322 

0.330 

0.338 

0.343 

0.362 

0.373 

10 

0.304 

0.312 

0.320 

0.328 

0.332 

0.350 

0.361 

12 

0.301 

0.308 

0.316 

0.324 

0.328 

0.346 

0.356 

H 

0.502 

0.514 

0.526 

0.539 

0.546 

0.577 

0.595 

H 

0.450 

0.461 

0.473 

0.484 

0.490 

0.517 

0.532 

i 

0.405 

0.415 

0.426 

0.436 

0.442 

0.466 

0.480 

i# 

0.369 

0.378 

0.387 

0.396 

0.401 

0.423 

0.435 

11A 

0.343 

0.352 

0.361 

0.370 

0.375 

0.397 

0.409 

2 

0.321 

0.329 

0.337 

0.345 

0.350 

0.369 

0.380 

m 

0.301 

0.309 

0.317 

0.324 

0,330 

0.348 

0.358 

3 

0.286 

0.293 

0.301 

0.308 

0.313 

0.330 

0.340 

1H 

VA 

0.274 

0.281 

0.288 

0.295 

0.300 

0.316 

0.326 

4 

0.267 

0.273 

0.280 

0.287 

0.291 

0.307 

0.317 

4K 

0.259 

0.266 

0.272 

0.279 

0.283 

0.299 

0,308 

5 

0.253 

0.260 

0.266 

0.272 

0.276 

0.291 

0.300 

6 

0.247 

0.253 

0.260 

0.266 

0.269 

0.284 

0.293 

8 

0.234 

0.240 

0.246 

0.252 

0.255 

0.270 

0.279 

10 

0,223 

0.229 

0.235 

0.241 

0.245 

0.258 

0.266 

12 

0.221 

0.227 

0.232 

0.238 

0.241 

0.255 

0.263 

1A 

0.442 

0.453 

0.464 

0.475 

0.481 

0.508 

0.523 

H 

0.392 

0.402 

0.412 

0.422 

0.428 

0.452 

0.465 

0.352 

0.360 

0.369 

0.378 

0.383 

0.405 

0.417 

itf 

0.319 

0.327 

0.335 

0.343 

0.348 

0.367 

0.379 

IK 

0.297 

0.304 

0.311 

0.319 

0.323 

0,341 

0.352 

2 

0.274 

0.280 

0.287 

0,294 

0.298 

0.314 

0.324 

2^ 

0.256 

0.262 

0.269 

0,275 

0.279 

0.293 

0.302 

3 

0,243 

0.249 

0.254 

0.260 

0.264 

0.277 

0.285 

2 

3H 

0.231 

0.236 

0.242 

0.248 

0.251 

0.265 

0.273 

4 

0.223 

0.228 

0.234 

0.240 

0.243 

0.257 

0.265 

4^ 

0.216 

0.222 

0.227 

0.233 

0.236 

0.249 

0.256 

5 

0.210 

0.215 

0.220 

0.225 

0,228 

0.241 

0.248 

6 

0.203 

0.208 

0.213 

0.218 

0.221 

0.233 

0.240 

8 

0491 

0.196 

0.201 

0.206 

0.209 

0.220 

0.227 

10 

0.182 

0.187 

0.192 

0.196 

0.199 

0.210 

0.215 

12 

0,178 

0.183 

0.187 

0.192 

0.195 

0.205 

0.210 

628 


CHAPTER  36 — INSULATION  OF  PIPING 


TABLE  9.    COEFFICIENTS  OF  TRANSMISSION  ( U)  FOR  PIPES  INSULATED  WITH  LAMINATED 
ASBESTOS  TYPE  INSULATION  (APPROXIMATELY  20  LAMINATIONS  PER  INCH  THICKNESS) 

These  coefficients  are  expressed  in  Btu  per  hour  per  square  foot  of  pipe  surface  per  degree 
Fahrenheit  difference  in  temperature  between  pipe  and  surrounding  still  air  at  70  F 


HOT  WATER 

STEAM 

THICKNESS 

NOMINAL, 

OP 

INSULATION 

PIPE 

SIZE 

120  F 

150  F 

180  F 

210  F 

227.1  F 
(5Lb) 

297.7  F 
(50  Lb) 

337.9  F 
(100  Lb) 

(INCHES) 

(INCHES) 

TEMPERATURE  DIFFERENCE 

50  F 

80  F 

110  F 

140  F 

157.1  F 

227.7  F 

267.9  F 

1A 

0.910 

0.925 

0.940 

0.956 

0.964 

1.001 

1.022 

H 

0.823 

0.836 

0.850 

0.863 

0.871 

0.902 

0.921 

i 

0.748 

0.760 

0.773 

0.785 

0.792 

0.823 

0.840 

1J4 

0.686 

0.698 

0.710 

0.721 

0.728 

0.756 

0.771 

VA 

0.649 

0.659 

0.671 

0.682 

0.688 

0.716 

0.731 

2 

0.610 

0.620 

0.630 

0.640 

0.647 

0.671 

0.685 

m 

0.581 

0.590 

0.600 

0.609 

0.615 

0.638 

0.651 

3 

0.558 

0.567 

0.576 

0.585 

0.591 

0.613 

0.626 

1 

3H 

0.539 

0.548 

0.557 

0.566 

0.571 

0.592 

0.604 

4 

0.524 

0.532 

0.541 

0.551 

0.556 

0.577 

0.589 

4H 

0.514 

0.522 

0.530 

0.539 

0.544 

0.564 

0.575 

0.503 

0.511 

0.519 

0.528 

0.533 

0.553 

0.565 

6 

0.492 

0.500 

0.509 

0.517 

0.522 

0.542 

0.553 

8 

0.473 

0.480 

0.488 

0.497 

0.502 

0,521 

0.532 

10 

0.458 

0.465 

0.473 

0.481 

0.485 

0.504 

0.514 

12 

0.452 

0.459 

0.467 

0.475 

0.478 

0.497 

0.507 

1A 

0.755 

0.767 

0.780 

0.793 

0.800 

0.831 

0.848 

H 

0.674 

0.685 

0.697 

0.708 

0.715 

0.743 

0.759 

i 

0.607 

0.618 

0.628 

0.639 

0.645 

0.670 

0.684 

U4 

0.553 

0.562 

0.572 

0.581 

0.587 

0.610 

0.622 

\1A 

0,517 

0.527 

0.536 

0.545 

0.550 

0.572 

0.584 

2 

0.481 

0.490 

0.499 

0.508 

0.513 

0.535 

0.547 

m 

0.453 

0,460 

0.469 

0.477 

0.481 

0.500 

0.511 

3 

0.429 

0.436 

0.444 

0.452 

0.456 

0.475 

0.485- 

1M 

3H 

0.412 

0.419 

0,427 

0.434 

0.438 

0.456 

0.465 

4 

0.400 

0.407 

0.415 

0.422 

0.426 

0.443 

0.453 

4M 

0.390 

0.396 

0.402 

0.409 

0.413 

0.429 

0.437 

5 

0.380 

0.386 

0,393 

0.400 

0.403 

0.418 

0.427 

6 

0.369 

0.375 

0.382 

0.389 

0,392 

0.408 

0.417 

8 

0.351 

0.358 

0.364 

0.370 

0,374 

0.388 

0.397 

10 

0,337 

0,344 

0,350 

0.356 

0.359 

0.373 

0.382 

12 

0.332 

0.338 

0.344 

0.350 

0.353 

0.367 

0,375 

H 

0.664 

0.675 

0.687 

0.698 

0.704 

0.732 

0.747 

H 

0.591 

0.601 

0,611 

0.621 

0.627 

0.652 

0.665 

i 

0.529 

0.538 

0.547 

0.557 

0.562 

0.584 

0.597 

134 

0.480 

0.488 

0.497 

0.505 

0.510 

0.529 

0.540- 

IK 

0.445 

0.453 

0.462 

0.470 

0.475 

0.494 

0.504 

2 

0,412 

0.420 

0.427 

0.434 

0.438 

0,455 

0.464 

2H 

0.385 

0.392 

0.398 

0.405 

0,409 

0.425 

0.434 

3 

0.364 

0.370 

0.376 

0.382 

0.385 

0.400 

0.408 

2 

3H 

0.346 

0.352 

0.358 

0.365 

0.368 

0.382 

0.390< 

4 

0.336 

0.342 

0.348 

0.354 

0.357 

0.371 

0.378 

4H 

0.325 

0.332 

0.338 

0.343 

0.346 

0.360 

0.367 

0.316 

0.322 

0.327 

0.333 

0,336 

0.349 

0.356- 

6 

0.306 

0.312 

0.317 

0,323 

0.326 

0.338 

0.345 

B 

0.288 

0.293 

0.298 

0.303 

0.306 

0.317 

•0.324 

10 

0.275 

0.279 

0.284 

0.289 

0.292 

0.302 

0.308 

12 

0.269 

0.274 

0.278 

0.283 

0.286 

0.296 

0.302 

629 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


TABLE  10.    COEFFICIENTS  OF  TRANSMISSION  (U)  FOR  PIPES  INSULATED 
WITH  ROCK  WOOL  TYPE  INSULATION 

These  coefficients  are  expressed  in  Btu  per  hour  per  square  foot  of  pipe  surface  per  degree 
Fahrenheit  difference  in  temperature  between  pipe  and  surrounding  still  air  at  70  F 


HOT  WATER 

STEAM 

THICKNESS 

NOMINAL 

OF 

INSIFLATION 

PIPE 

SIZE 

120  F 

1SOF 

180F 

210  F 

227.1  F 
(SLb) 

297.7  F 
(50  Lb) 

337.9  F 
(100  Lb) 

(INCHES) 

(INCHES) 

TEMPERATURE  DIFFERENCE 

50  F 

80  F 

110  F 

140  F 

157.1  F 

227.7  F 

2673F 

Ji 

0.631 

0.644 

0.658 

0.672 

0.680 

0.712 

0.730 

H 

0.569 

0.581 

0.593 

0.606 

0.613 

0.642 

0.659 

0.518 

0.529 

0.541 

0.552 

0.559 

0.585 

0.600 

1J4 

0.476 

0.486 

0.497 

0.507 

0.513 

0.537 

0.551 

1H 

0.450 

0.460 

0.470 

0.480 

0.485 

0.508 

0.522 

2 

0.422 

0.431 

0.441 

0.450 

0.456 

0.478 

0.490 

2H 

0.402 

0.411 

0.420 

0.428 

0.434 

0.455 

0.466 

3 

0.385 

0.394 

0.402 

0.411 

0.415 

0.435 

0.446 

1 

3}i 

0.373 

0.381 

0.389 

0.398 

0.402 

0.421 

0.432 

4 

0.363 

0.371 

0.379 

0.387 

0.392 

0.411 

0.422 

VA 

0.355 

0.363 

0.371 

0.379 

0.383 

0.402 

0.413 

0.348 

0.356 

0.364 

0.371 

0.376 

0.394 

0.404 

6 

0.341 

0.348 

0.356 

0.363 

0.368 

0.386 

0.396 

8 

0.327 

0.335 

0.342 

0.349 

0.353 

0.372 

0.381 

10 

0.317 

0.324 

0.331 

0.338 

0.343 

0.360 

0.369 

12 

0.313 

0.320 

0.327 

0.334 

0.338 

0.355 

0.364 

y2 

0.523 

0.534 

0.545 

0.556 

0.563 

0.590 

0.606 

u 

0.468 

0.477 

0.487 

0.497 

0.503 

0,528 

0.542 

i 

0.421 

0.430 

0.440 

0.449 

0.455 

0.477 

0.490 

IK 

0.383 

0.391 

0.399 

0.407 

0.412 

0.433 

0.444 

i« 

0.359 

0.366 

0.375 

0.383 

0.387 

0.407 

0.419 

2 

0.333 

0.340 

0.348 

0.356 

0.360 

0.378 

0.389 

2J^ 

0.314 

0.320 

0.327 

0.335 

0.339 

0.355 

0.365 

3 

0.296 

0.302 

0.310 

0,317 

0.321 

0.337 

0.347 

1H 

m 

0.286 

0.291 

0.298 

0.304 

0.307 

0.323 

0.332 

4 

0.278 

0.284 

0.290 

0.296 

0.300 

0.315 

0.323 

4^ 

0.270 

0.276 

0.282 

0,287 

0.291 

0.305 

0,313 

5 

0.263 

0.269 

0.275 

0.280 

0.284 

0.298 

0.305 

6 

0.257 

0.262 

0.267 

0.273 

0.277 

0.290 

0.297 

8 

0.244 

0.249 

0.254 

0.260 

0.263 

0.276 

0.283 

10 

0.235 

0.240 

0.245 

0.250 

0.253 

0.265 

0.272 

12 

0.230 

0.234 

0.239 

0.245 

0.247 

0.260 

0.267 

« 

0.461 

0.471 

0.481 

0.491 

0,496 

0.520 

0.534 

« 

0.409 

0.418 

0.427 

0.436 

0,441 

0.463 

0.475 

0.366 

0.374 

0.382 

0.390 

0.395 

0.415 

0.427 

IK 

0.333 

0.340 

0.347 

0,355 

0.359 

0.377 

0.387 

1« 

0.310 

0.316 

0.323 

0.330 

0.334 

0.351 

0.360 

2 

0.286 

0.292 

0.298 

0.304 

0.308 

0.323 

0.331 

2M 

0.268 

0.274 

0.279 

0.285 

0.289 

0.302 

0,310 

3 

0.252 

0.257 

0.262 

0.268 

0.272 

0.284 

0.292 

2 

m 

0.241 

0.246 

0.251 

0.257 

0.260 

0.272 

0,280 

4 

0.232 

0.237 

0.242 

0.247 

0.250 

0.262 

0,269 

4H 

0.225 

0.230 

0.235 

0.240 

0.243 

0.255 

0,262 

5 

0.218 

0.223 

0.228 

0.233 

0.236 

0.247 

0.253 

6 

0.213 

0,217 

0.221 

0.226 

0.228 

0.239 

0.245 

8 

0.200 

0.204 

0.208 

0.213 

0,215 

0.225 

0.231 

10 

0.189 

0.193 

0.197 

0.201 

0.204 

0,214 

0,220 

12 

0.185 

0,190 

0.194 

0.198 

0.200 

0.210 

0,216 

630 


CHAPTER  36 — INSULATION  OF  PIPING 


25  F;  the  thickness  of  insulation  for  brine  is  approximately  2J^  in.  where 
the  temperature  ranges  from  0  deg  to  25  F;  and  the  thickness  of  insulation 
where  the  brine  temperature  ranges  from  —30  F  to  zero  degrees  is  ap- 
proximately 4  in. 

Insulation  To  Prevent  Freezing 

If  the  surrounding  air  temperature  remains  sufficiently  low  for  an 
ample  period  of  time,  insulation  cannot  prevent  the  freezing  of  still  water, 
or  of  water  flowing  at  such  a  velocity  that  the  quantity  of  heat  carried  in 
the  water  is  not  sufficient  to  take  care  of  the  heat  losses  which  will  result 
and  cause  the  temperature  of  the  water  to  be  lowered  to  the  freezing 
point.  Insulation  can  materially  prolong  the  time  required  for  the  water 
to  give  up  its  heat,  and  if  the  velocity  of  the  water  flowing  in  the  pipe  is 
maintained  at  a  sufficiently  high  rate,  freezing  may  be  prevented. 
^  Table  11  may  be  used  for  making  estimates  of  the  thickness  of  insula- 
tion necessary  to  take  care  of  still  water  in  pipes  at  various  water  and 
surrounding  air  temperature  conditions.  Because  of  the  damage  and 
service  interruptions  which  may  result  from  frozen  water  in  pipes,  it  is 
essential  that  the  most  efficient  insulation  be  utilized.  This  table  is 
based  on  the  use  of  hair  felt  or  cork,  having  a  conductivity  of  0.30.  The 
initial  water  temperature  is  assumed  to  be  10  deg  above,  and  the  sur- 
rounding air  temperature  50  deg  below  the  freezing  point  of  water  (tem- 
perature difference,  60  F). 

The  last  column  of  Table  11  gives  the  minimum  quantity  of  water  at 
initial  temperature  of  42  F  which  should  be  supplied  every  hour  for  each 
linear  foot  of  pipe,  in  order  to  prevent  the  temperature  of  the  water  from 
being  lowered  to  the  freezing  point.  The  weights  given  in  this  column 
should  be  multiplied  by  the  total  length  of  the  exposed  pipe  line  expressed 
in  feet.  As  an  additional  factor  of  safety,  and  in  order  to  provide  against 
temporary  reductions  in  flow  occasioned  by  reduced  pressure,  it  is 
advisable  to  double  the  rates  of  flow  listed  in  the  table.  It  must  be 
emphasized  that  the  flow  rates  and  periods  of  time  designated  apply  only 
for  the  conditions  stated.  To  estimate  for  other  service  conditions  the 
following  method  of  procedure  may  be  used. 

If  water  enters  the  pipe  at  52  F  instead  of  42  F,  the  time  required  to 
cool  it  to  the  freezing  point  will  be  prolonged  to  twice  that  given  in  the 
table,  or  the  rate  of  flow  of  water  may  be  reduced  so  that  the  quantity 
required  will  be  one-half  that  shown  in  the  last  column  of  Table  11. 
However,  if  the  water  enters  the  pipe  at  34  F  it  will  be  cooled  to  32  F  in 
one-fifth  of  the  time  given  in  the  table.  It  will  then  be  necessary  to  in- 
crease the  rate  of  flow  so  that  five  times  the  specified  quantity  of  water 
will  have  to  be  supplied  in  order  to  prevent  freezing. 

If  the  minimum  air  temperature  is  —  38  F  (temperature  difference, 
80  F),  instead  of  —18  F,  the  time  required  to  cool  the  water  to  the 
freezing  point  will  be  60/80  of  the  time  given  in  the  table,  or  the  necessary 
quantity  of  water  to  be  supplied  will  be  80/60  of  that  given. 

In  making  calculations  to  arrive  at  the  values  given  in  Table  11,  the  loss 
of  heat  stored  in  the  insulation,  the  effect  of  a  varying  temperature  dif- 
ference due  to  the  cooling  of  pipe  and  water,  and  the  resistance  of  the 
outer  surface  of  the  insulation  to  the  transfer  of  heat  to  the  air  have  all 

681 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


TABLE  11. 


DATA  FOR  ESTIMATING  REQUIREMENTS  TO  PREVENT 
FREEZING  OF  WATER  IN  PIPES 


NOMINAL 

NUMBER  OF  HOURS 

WATER  REQUIRED  TO  FLOW 

PIPE 

TO  COOL 

TO  PREVENT  FREEZING, 

SIZE 

WATER  TO 

POUNDS  PER  LINEAR  FOOT  OF 

(INCHES) 

FREEZING  POINT 

PIPE  PER  HOUR 

Thickness  of  Insulation  in  Inches 

1 

2 

3 

1 

2 

3 

1A 

0.42 

0.50 

0.57 

0.54 

0.45 

0.40 

i 

0.83 

1.02 

1.16 

0.68 

0.55 

0.48 

1H 

1.40 

1.74 

2.02 

0.84 

0.68 

0.58 

2 

1.94 

2.48 

2.90 

0.95 

0.75 

0.64 

3 

3.25 

4.27 

5.08 

1.24 

0.94 

0.79 

4 

4.55 

6.02 

7.20 

1.47 

1.11 

0.93 

5 

5.92 

7.96 

9.69 

1.73 

1.29 

1.06 

6 

7.35 

9.88 

12.20 

1.98 

1.46 

1.19 

8 

10.05 

13.90 

17.25 

2.46 

1.78 

1.44 

10 

13.00 

18.10 

22.70 

2.96 

2.12 

1.70 

12 

15.80 

22.20 

28.10 

3.43 

2.46 

1.93 

been  neglected.  When  these  factors  enter  into  the  computations  it  is 
necessary  to  enlarge  the  factor  of  safety.  Also  as  stated,  the  time  shown 
in  the  table  is  that  required  to  lower  the  water  to  the  freezing  point.  A 
longer  period  would  be  required  to  freeze  the  water,  but  the  danger  point 
is  reached  when  freezing  starts.  The  flow  of  water  will  stop  and  the  entire 
line  will  be  in  danger  as  soon  as  the  water  freezes  across  the  section  of  the 
pipe  at  any  point. 

When  water  must  remain  stationary  longer  than  the  times  designated  in 
Table  11,  the  only  safe  way  to  insure  against  freezing  is  to  install  a  steam 
or  hot  water  line,  or  to  place  an  electric  resistance  heater  along  the  side  of 
the  exposed  water  line.  The  heating  system  and  the  water  line  are  then 
insulated  so  that  the  heat  losses  from  the  heating  system  are  not  exces- 
sive, and  the  heating  effect  is  concentrated  against  the  water  pipe  where 
it  is  needed.  For  this  form  of  protection  2  in.  of  an  efficient  insulation 
may  be  applied. 

Pipe  Sweating 

In  some  cases  the  prevention  of  condensation  rather  than  the  con- 
servation of  heat  is  the  governing  factor  in  determining  the  thickness  of 
insulation  required.  Fig.  2  may  be  used  for  determining  the  thickness  of 
any  material  of  known  conductivity  which  should  be  used  to  prevent  con- 
densation on  pipes  and  flat  metallic  surfaces.  The  surface  resistances  used 
for  calculating  the  family  of  curves  in  Fig.  2  are  based  on  the  results  of 
tests  made  on  canvas-covered  pipe  insulation  surfaces  at  Mellon  Institute. 
However,  it  has  been  found  that  the  resistance  for  asphaltic  and  roofing 
surfaces  is  practically  the  same  as  for  canvas  surfaces,  so  that  the  curves 
given  may  be  followed  with  no  alteration  for  surfaces  commonly  used. 

Moisture  will  be  deposited  on  a  surface  whenever  its  temperature  falls 
to  that  of  the  dew  point,  The  maximum  permissible  temperature  drop 
is  indicated  on  Fig.  2  at  the  point  where  the  guide  line  passes  through  the 
horizontal  scale  at  the  left  center  of  the  chart.  This  temperature  drop 

632 


CHAPTER  36 — INSULATION  OF  PIPING 


represents  the  difference  between  the  dry-bulb  temperature  and  the  dew- 
point  temperature  for  the  conditions  involved.  (See  discussion  of  con- 
densation in  Chapter  7.) 

The  rate  of  heat  loss  from  a  surface  maintained  at  constant  temperature 
is  greatly  increased  by  air  circulation  over  the  surface.  In  the  case  of 
well-insulated  surfaces  the  increases  in  losses  due  to  air  velocity  are  very 
small  as  compared  with  increases  shown  for  bare  surfaces,  because  of  the 


\     \       ^       \        \ 

80    70     60      50      4-0     30         20 


FIG.  2.   THICKNESS  OF  PIPE  INSULATION  TO  PREVENT  SWEATING* 

aSolve  problems  by  drawing  lines  as  indicated  by  dotted  line,  entering  chart  at  lower  left  hand  scale. 

fact  that  air  flowing  over  the  surface  of  the  insulation  can  increase  only 
the  rate  of  heat  transfer  from  surface  to  air,  and  cannot  change  the  internal 
resistance  to  heat  flow  inherent  in  the  insulation  itself.  The  maximum 
increase  in  loss  due  to  air  velocity  ranges  from  about  30  per  cent  in  the 
case  of  1-in.  thick  insulation,  to  about  10  per  cent  in  the  case  of  3-in.  thick 
insulation,  provided  that  the  insulation  is  thoroughly  sealed  so  that  air 
can  flow  only  over  the  surface. 

633 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


ECONOMICAL  THICKNESS  OF  INSULATION. 


HOURS  OPERATION  PER  YEAR 


(L,  B,  McMillan,  Bw.  N&tional  Dtit.  Heating  Axs'n*,  Vol.  18,  p,  131.) 

FIG,  3.  CHART  FOR  DETERMINING  ECONOMICAL  THICKNESS  OF  INSULATION 

634 


CHAPTER  36 — INSULATION  OF  PIPING 


If  the  conditions  are  such  that  the  air  may  circulate  through  cracks 
and  crevices  in  the  insulation,  the  increases  may  be  far  greater  than  those 
given.  Therefore,  it  is  essential  that  insulation  be  sealed  as  tightly 
as  possible.  Pipe  insulation  out-of-doors  should  be  provided  with  a 
waterproof  jacket,  and  other  outdoor  insulation  should  be  thoroughly 
weatherproof  ed . 

ECONOMICAL  THICKNESS  OF  PIPE  INSULATION 

Table  12  shows  the  thicknesses  of  insulation  which  ordinarily  are  used 
for  various  temperature  conditions.  Where  a  thorough  analysis  of 
economic  thickness  is  desired,  this  may  be  accomplished  through  the  use 
of  the  chart,  Fig.  3. 

The  dotted  line  on  the  chart  illustrates  its  use  in  solving  a  typical 
example.  In  using  the  chart,  start  with  the  scale  at  the  left  bottom  margin 
representing  the  given  number  of  hours  of  operation  per  year;  then 
proceed  vertically  to  the  line  representing  the  given  value  of  heat ;  thence 
horizontally,  to  the  right,  to  the  line  representing  the  given  temperature 
difference;  thence  vertically  to  the  line  representing  the  conductivity  of 

TABLE  12.    THICKNESSES  OF  INSULATION  ORDINARILY  USED  lNDOORSa 


THICKNESS  OF  INSULATION 

STEAM  PRESSURES 

STEAM  TEMPERATURES 

(Le  GAGE) 

DEGREES 

OR  CONDITIONS 

FAHRENHEIT 

Pipes  Larger 
nrhnn  i  Tn 

Pipes 
2  In.  to 

Pipes 
HIn. 

4  In. 

to  1H  In. 

Oto25 

212  to  267 

lin. 

lin. 

lin. 

25  to  100 

267  to  338 

l^in. 

1  in. 

1  in. 

100  to  200 

338  to  388 

2  in. 

IJ^in. 

lin. 

Low  Superheat 

388  to  500 

2^  in. 

2  in. 

1^2  in. 

Medium  Superheat 

500  to  600 

3  in. 

2%  in. 

2  in. 

High  Superheat 

600  to  700 

3^  in. 

3  in. 

2  in. 

aAU  piping  located  outdoors  or  exposed  to  weather  is  ordinarily  insulated  to  a  thickness  M  in.  greater 
than  shown  in  this  table,  and  covered  with  a  waterproof  jacket. 

the  given  material ;  thence  horizontally,  to  the  left,  to  the  line  representing 
the  given  discount  on  that  material;  thence  vertically  to  the  curve 
representing  the  required  per  cent  return  on  the  investment;  thence 
horizontally,  to  the  right,  to  the  curve  representing  the  given  pipe  size; 
thence  vertically  to  the  scale  at  the  top  right  margin  where  the  economical 
thickness  may  be  read  off  directly.  The  dotted  line  on  the  chart  illustrates 
its  use  in  solving  a  typical  example. 

Underground  Insulation 

Underground  steam  distribution  lines  are  carried  in  protective  struc- 
tures of  various  types,  sizes  and  shapes.  (See  Chapter  37.)  Detailed 
data  on  commonly  used  forms  of  tunnels  and  conduit  systems  have  been 
published  by  the  National  District  Heating  Association*. 

Pipes  in  tunnels  are  covered  with  sectional  insulation  to  provide 
maximum  thermal  efficiency  and  are  also  finished  with  good  mechanical 


^Handbook  ctf  the  National  District  Seating  Association,  Second  Edition,  1932. 

635 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

protection  in  the  form  of  metal  or  waterproofing  membrane  outer 
jackets.  Conduit  systems  are  in  more  general  use  than  tunnels.  Pipes 
carried  in  conduits  may  be  insulated  with  sectional  insulation;  however, 
the  more  usual  practice  is  to  fill  the  entire  section  of  the  conduit  around 
the  pipes  with  high  quality,  loose  insulating  material.  The  insulation 
must  be  kept  dry  at  all  times,  and  for  this  purpose  effective  waterproofing 
membranes  enclose  the  insulation.  A  drainage  system  is  also  provided 
to  divert  water  which  may  tend  to  enter  the  conduit. 

The  economical  thickness  of  insulation  for  underground  work  is  dif- 
ficult of  accurate  determination  due  to  the  many  variables  which  have  to 
be  considered.  As  a  result  of  theories  developed  by  J.  R.  Allen3,  together 
with  experimental  data  presented  by  others,  the  usual  endeavor  is  to 
secure  not  less  than  90  per  cent  efficiency  for  underground  piping.  Table 
13  can  be  used  as  a  guide  in  arriving  at  the  minimum  thickness  of  loose 
insulation  fills  to  use  for  laying  out  conduit  systems.  Other  factors  such 
as  the  number  of  pipes  and  their  combination  of  sizes,  as  well  as  the 
standard  conduit  sizes,  are  primary  controlling  factors  in  the  amount  and 
thickness  of  insulation  for  use. 

When  sectional  insulation  is  applied  to  lines  in  tunnels  or  conduits, 
usual  practice  is  to  apply  the  most  efficient  materials  J^  in.  less  in  thick- 
ness than  that  determined  by  the  use  of  Fig.  3.  Fig.  3  is  based  on  con- 
ditions of  insulation  exposed  to  the  air,  whereas  normal  ground  tempera- 
ture is  substituted  for  air  temperature  in  determining  the  temperature 
difference  for  use  with  the  chart  when  applying  it  for  underground  pipe 
line  estimates. 

TABLE  13.    THICKNESS  OF  LOOSE  INSULATION  FOR  USE  AS 
FILL  IN  UNDERGROUND  CONDUIT  SYSTEMS 


MINIMUM  THICKNESS  OF  INSULATION  IN  INCHES 

MINIMUM 

STEAM 

STEAM 

DISTANCE 

PRESSURES 
(La  GAGE) 

TEMPERATURES 

DEGREES 

STEAM  LINES 

RETURN  LINES 

BETWEEN 
STEAM 

OR  CONDITIONS 

FAHRENHEIT 

Pipes  Less 
than  4  In. 

Pipes  4  In. 
to  10  In. 

Pipes  Larger 
than  12  In. 

Pipes  Less 
than  4  In. 

Pipes  4  In. 
and  Larger 

AND 

RETURN 

Hot  Water, 

or  0  to  25 

212  to  267 

1  3/i-j 

2 

2j/]2 

1M 

1J"12 

1 

25  to  125 

267  to  352 

2 

2M 

3 

!}<£ 

1M 

iM 

Above  125,  or 

superheat 

352  to  500 

W* 

3 

m 

1M 

VA 

m 

•Theory  of  Heat  Losses  from  Pipes  Buried  in  the  Ground,  by  J,  R,  Allen  (A.S.H.V.E.  TRANSACTIONS. 
Vol.  26,  1920). 


PROBLEMS  IN  PRACTICE 

1  •  Compute  the  total  annual  heat  loss  from  165  ft  of  2-in.  bare  pipe  in  service 
4000  hours  per  year.  The  pipe  is  carrying  steam  at  10  lb  pressure  and  is  exposed 
to  an  average  air  temperature  of  70  F. 

The  pipe  temperature  is  taken  as  the  steam  temperature,  which  is  239.4  F,  obtained 
from  Table  7,  Chapter  1.  The  temperature  difference  between  the  pipe  and  air  «=  239.4 
—  70  «  169.4  deg.  By  interpolation  of  Table  1  between  temperature  differences  of 
157.1  F  and  227.7  F,  the  heat  loss  from  a  2-in.  pipe  at  a  temperature  difference  of  169.4 

636 


CHAPTER  36 — INSULATION  OF  PIPING 


deg  is  found  to  be  1.677  Btu  per  hour  per  linear  foot  per  degree  temperature  difference. 
The  total  annual  heat  loss  from  the  entire  =  1.677  X  169.4  X  165  (linear  feet)  X  4000 
(hours)  =  188,000,000  Btu. 

2  •  Coal  costing  $11.50  per  ton  and  having  a  calorific  value  of  13,000  Btu  per 
pound  is  being  burned  in  the  furnace  supplying  steam  to  the  pipe  line  given  in 
Question  1.     If  the  system  is  operating  at  an  over-all  efficiency  of  55  per  cent 
determine  the  monetary  value  of  the  annual  heat  loss  from  the  line. 

The  cost  of  heat  per  1  million  Btu  supplied  to  the  system  =  1,000,000  X  11.5  (dollars) 
-T-  13,000  (Btu)  X  2000  (Ib)  X  0.55  (efficiency)  =  $0.804.  The  total  cost  of  heat 
lost  per  year  =  0.804  X  188  (million  Btu)  =  $151. 15.4 

3  •  If  the  steam  line  given  in  Question  1  is  covered  with  1-in.  thick  85  per  cent 
magnesia,  determine  the  resulting  total  annual  heat  loss  through  the  insula- 
tion.    Also  compute  the  monetary  value  of  the  annual  saving  and  the  per- 
centage of  saving  over  the  heat  loss  from  the  bare  pipe. 

By  interpolation  of  Table  5  between  temperature  differences  of  157.1  F  and  227.7  F,  the 
coefficient  of  transmission  for  1-in.  magnesia  on  a  2-in.  pipe  is  found  to  be  0.525  Btu  per 
hour  per  square  foot  of  pipe  surface  per  degree  temperature  difference  at  a  temperature 
difference  of  169.4  deg.  The  total  hourly  loss  per  square  foot  of  insulated  pipe  will  then 
be  0.525  X  169.4  «  89.04  Btu.  From  Table  2  the  area  per  linear  foot  of  2-in.  pipe  is 
found  to  be  0.622  sq  ft.  The  total  annual  loss  through  the  insulation  =  89.04  X  0.622 
X  165  (linear  feet)  X  4000  (hours)  =  36,550,000  Btu.  The  annual  bare  pipe  loss  as 
determined  in  the  solution  of  Question  1  was  found  to  be  188,000,000  Btu.  The  saving 
due  to  insulation  is  then  188,000,000  —  36,550,000  =  151,350,000  Btu  per  year. 

From  the  solution  of  Question  2  it  was  found  that  the  heat  supplied  to  the  system  cost 
$0.804  per  million  Btu ;  therefore,  the  monetary  value  of  the  saving  =  0.804  (dollars) 
X  151.35  (million  Btu)  =  $121.69,  or  81.2  per  cent  of  the  cost  when  using  uninsulated 
pipe. 

4  •  The  manufacturer's  list  price  for  85  per  cent  magnesia  insulation  is  $0.36 
per  linear  foot  for  1-in.  (standard  thick)  material  to  cover  a  2-in.  pipe.    De- 
termine the  period  of  time  required  for  the  saving  found  in  Question  3  to  pay 
for  the  cost  of  the  insulation  if  it  can  be  purchased  and  applied  at  80  per  cent 
of  list  price  (20  per  cent  discount). 

The  applied  cost  of  insulation  =  165  (linear  feet)  X  0.36  (dollars)  X  0.80  (net) 
—  47.52.  Since  the  annual  saving  as  found  in  Question  3  amounts  to  $121.69,  the  in- 
sulation will  pay  for  its  cost  in  47.52  -f-  121.69  —  0,3905  years;  in  other  words,  the  cost 
will  be  repaid  2.56  times  by  the  saving  obtained  in  one  heating  season. 

5  •  The  conductivity  of  magnesia  insulation  is  0.455  at  the  mean  temperature 
which  will  result  under  the  conditions  of  Question  3.     Estimate  the  most 
economical  thickness  of  magnesia  for  application  on  the  pipe  when  operating 
under  the  conditions  which  are  given  in  the  foregoing  problems  and  when  a 
20  per  cent  return  is  required  on  the  investment  for  insulation. 

Use  chart  given  in  Fig.  3.  Begin  at  the  left  bottom  margin  and  proceed  successively  as 
shown  by  the  dotted  line  example  to  the  following  essential  data  which  are  collected  from 
the  problems  previously  given; 

4000  hours  operation  per  year, 

$0.804  value  of  heat,  dollars  per  million  Btu. 

169.4  deg  temperature  difference. 

0.455  conductivity  of  insulation, 

20  per  cent  discount  from  list,  cost  of  insulation, 

20  per  cent  fixed  charges,  return  on  investment. 

2-in.  pipe  size. 

Solution  of  the  problem  by  use  of  Fig.  3  results  in  a  required  thickness  of  approximately 


4  A  closely  approximate  solution  of  this  problem  may  be  quickly  made  by  use  of  the  estimating  chart  given 
In  Fig,  1. 

637 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

1.05  in.  The  nearest  commercial  thickness  procurable  is  standard  thick  (1^2  in.) 
magnesia. 

(It  is  of  interest  to  note  that  the  use  of  Fig.  3  will  generally  result  in  solutions  which,  for  all  practical  pur- 
poses, agree  closely  with  the  specifications  for  thicknesses  given  in  Table  12.) 

6  •  Determine  the  minimum  thickness  of  wool  felt  insulation  having  a  con- 
ductivity of  0.30  necessary  to  prevent  condensation  of  moisture  on  a  4-in.  pipe 
carrying  cold  water  at  a  temperature  of  40  F  when  the  surrounding  air  reaches 
maximum  conditions  of  90  F  with  a  relative  humidity  of  90  per  cent. 

The  difference  between  the  temperature  of  the  pipe  and  the  surrounding  air  is  90  —  40 
=  50  deg.  For  quick  estimating  purposes  use  the  chart  given  in  Fig.  2.  Enter  this  chart 
at  the  lower  left  margin  on  the  90  per  cent  relative  humidity  line  and  proceed  horizontally 
to  the  right  to  intersect  the  90  deg  air  temperature  line.  Project  a  line  up  to  the  50  deg 
temperature  difference  line,  and  then  horizontally  to  the  right  to  the  intersection  with  the 
4-in.  pipe  size  line.  From  this  point  proceed  down  to  intersect  the  0.30  line  which  denotes 
the  conductivity  of  the  insulation.  Directly  opposite  this  point  of  intersection  the  correct 
thickness  of  insulation  is  read  from  the  scale  on  the  lower  right  margin.  This  chart 
solution  denotes  that  wool  felt  2.4-in.  thick  is  sufficient  to  prevent  condensation.  The 
nearest  commercial  thickness  procurable  is  2J^  in. 

For  prevention  of  condensation  as  well  as  for  protection  against  freezing,  if  the  thickness 
determined  theoretically  cannot  be  had,  it  is  better  to  apply  the  next  greater  thickness 
procurable  rather  than  to  use  any  lesser  thickness  because  an  additional  factor  of  safety 
is  thus  obtained. 

7  •  A  3-in.  pipe  covered  with  2  in.  of  hair  felt  insulation  carries  water  out-of- 
doors.    Weather  Bureau  records  for  the  locality  denote  that  a  minimum,  out- 
door temperature  of  zero  F  may  be  expected  to  prevail  for  a  period  not  to  exceed 
10  hours.    By  use  of  Table  11  determine  what  degree  of  protection  is  provided 
against  freezing  if  the  water  is  stationary  in  the  line  for  the  10  hour  period. 

Table  11  denotes  that  4.27  hours  are  required  to  lower  water  temperature  from  42  F  to 
the  freezing  point  (a  10  F  drop)  when  the  initial  temperature  difference  between  the 
water  and  air  is  60  F.  The  temperature  drop  in  the  example  is  45  —  32  -  13  F  and  the 
temperature  difference  is  45  —  0  -45  F.  The  time  required  to  lower  the  water  from 
45  F  to  the  freezing  point  will  therefore  be  4.27  X  1Ko  X  6Ms  =  7.4  hours. 

It  is  evident  that  insufficient  protection  is  provided  to  prevent  freezing  if  a  temperature 
of  zero  F  prevails  outside  for  any  period  longer  than  7  hours  24  min. 

8  •  With  data  given  in  Question  7  and  its  solution,  determine  the  minimum 
flow  of  water  which  must  be  maintained  to  prevent  freezing  if  the  length  of 
the  water  line  is  85  ft. 

From  Table  11  it  is  seen  that  a  flow  of  0.94  Ib  of  water  per  linear  foot  per  hour  is  suf- 
ficient to  prevent  a  drop  in  the  water  temperature  from  42  F  to  the  freezing  point  when 
the  temperature  difference  between  air  and  water  is  60  F.  With  the  conditions  stated,  a 
flow  of  0.94  X  1%3  X  4%o  »  0.542  Ib  of  water  per  linear  foot  of  pipe  line  per  hour,  or 
85  X  0.642  «  46.07  Ib  of  water  per  hour  must  flow  through  the  line  in  order  to  prevent 
freezing. 


638 


Chapter  37 

DISTRICT  HEATING 

Underground  Steam  Piping,  Selection  of  Pipe  Sizes,  Provision  for 
Expansion9  Capacity  of  Returns  with  Various  Grades,  Pipe  Con- 
duits, Pipe  Tunnels,  Service  Connections,  Steam  per  Square  Foot 
of  Heating  Surface,  Fluid  Meters  and  Metering,  Rates 

THOSE  phases  of  district  heating  which  frequently  fall  within  the 
province  of  the  heating  engineer  are  outlined  here  with  data  and 
information  for  solving  incidental  problems  in  connection  with  institutions 
and  factories  and  for  the  design  of  heating  systems  for  buildings  which  are 
to  be  supplied  with  purchased  steam.  A  complete  district  heating  instal- 
lation should  not  be  attempted  without  a  thorough  study  of  the  entire 
problem  by  men  competent  and  experienced  in  that  industry. 

UNDERGROUND  STEAM  PIPING 

The  methods  used  in  district  heating  work  for  the  distribution  of  steam 
are  applicable  to  any  problem  involving  the  supply  of  steam  to  a  group  of 
buildings.  The  first  step  is  to  establish  the  route  of  the  pipes,  and  in  this 
matter  the  local  conditions  so  fully  control  the  layout  that  little  can  be 
said  regarding  it. 

Having  established  the  route  of  the  pipes,  the  next  step  is  to  calculate 
the  pipe  sizes.  In  district  heating  work  it  is  common  practice  to  design 
the  piping  system  on  the  basis  of  pressure  drop.  The  initial  pressure  and 
the  minimum  permissible  terminal  pressure  are  specified  and  the  pipe 
sizes  are  so  chosen  that  the  required  amount  of  steam,  with  suitable 
allowances  for  future  increases,  will  be  transmitted  without  exceeding 
this  pressure  drop.  The  steam  velocity  is  therefore  almost  disregarded 
and  may  reach  a  very  high  figure.  Velocities  of  35,000  fpm  are  not  con- 
sidered high.  By  the  use  of  this  method  the  pipe  sizes  are  kept  to  a 
minimum  with  consequent  savings  in  investment. 

The  steam  flowing  through  any  section  of  the  piping  can  be  computed 
from  a  study  of  the  requirements  of  the  several  buildings  served.  In 
general  a  condensation  rate  of  0.25  Ib  per  hour  per  square  foot  of  equiva- 
lent heating  surface  is  a  safe  figure.  This  allows  for  line  condensation 
which,  however,  is  a  small  part  of  the  total  at  times  of  maximum  load. 
Any  unusual  requirements  such  as  those  for  process  steam  should  be 
individually  calculated. 

The  steam  requirements  for  water  heating  should  be  taken  into  account, 
but  in  most  types  of  buildings  this  load  will  be  relatively  small  compared 
with  the  heating  load  and  will  seldom  occur  at  the  time  of  the  heating 

639 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

peak.  Unusual  features  such  as  large  heaters  for  swimming  pools  should 
not  be  overlooked. 

The  pressure  at  which  the  steam  is  to  be  distributed  will  depend,  in 
part,  upon  whether  or  not  it  has  been  passed  through  electrical  generating 
units.  If  it  has,  the  pressure  will  be  considerably  lower  than  if  live  steam, 
•direct  from  the  boilers,  is  used.  The  advantages  of  low  pressure  distribu- 
tion (2  to  30  Ib  per  square  inch)  are  (1)  smaller  heat  loss  from  the  pipes, 
(2)  less  trouble  with  traps  and  valves,  and  (3)  simpler  problems  in  pressure 
reduction  at  the  buildings.  With  distribution  pressures  not  exceeding 
40  Ib  per  square  inch  there  is  little  danger  even  if  the  full  distribution 
pressure  should  build  up  in  the  radiators  through  the  faulty  operation  of 
a  reducing  valve;  but  with  pressures  higher  than  this  a  second  reducing 
valve  or  some  form  of  emergency  relief  is  usually  desirable  to  prevent 
excessive  pressures  in  the  radiators.  The  advantages  of  high  pressure 
distribution  are  (1)  smaller  pipe  sizes  and  (2)  greater  adaptability  of  the 
steam  to  various  operations  other  than  building  heating. 

The  different  kinds  of  apparatus  which  frequently  must  be  served 
require  various  minimum  pressures.  Kitchen  equipment  requires  from 
5  to  15  Ib  per  square  inch,  the  higher  pressures  being  necessary  for 
apparatus  in  which  water  is  boiled,  such  as  stock  kettles  and  coffee  urns. 
An  increased  amount  of  heating  surface,  which  is  easily  obtained  in  some 
kinds  of  apparatus,  results  in  quicker  and  more  satisfactory  operation  at 
low  pressures.  For  laundry  equipment,  particularly  the  mangle,  a  pres- 
sure of  75  Ib  per  square  inch  is  usually  demanded  although  30  Ib  per  square 
inch  is  sufficient  if  the  mangle  is  equipped  with  a  large  number  of  rolls  and 
if  a  slow  rate  of  operation  is  permissible.  Pressing  machines  and  hospital 
-sterilizers  require  about  50  Ib  per  square  inch. 

PIPE  SIZES 

The  lengths  of  pipe,  steam  quantities,  and  initial  and  terminal  pressures 
having  been  chosen,  the  pipe  sizes  can  readily  be  calculated  by  means  of 
the  Unwin  pressure  drop  formula.  This  formula,  which  gives  pressure 
-drops  slightly  larger  than  actual  test  results,  is  as  follows: 


0.0001306  W*L  (*  +  ~f  ) 


(1) 


where 

P  =»  pressure  drop,  pounds  per  square  inch. 
W  =  weight  of  steam  flowing,  pounds  per  minute. 
L  «s  length  of  pipe,  feet. 
d  as  inside  diameter  of  pipe,  inches. 
y  as  average  density  of  steam,  pounds  per  cubic  foot. 

This  formula  is  similar  to  the  Babcock  formula  given  in  Chapter  32. 

Information  on  provision  for  expansion  will  be  found  in  Chapters 
32  and  34. 

In  general,  return  lines  when  installed  follow  the  contour  of  the  land, 
and  Table  1  gives  sizes  of  return  pipes  for  various  grades.  It  is  evident 
that  at  points  where  the  grade  is  great,  smaller  pipes  can  be  installed. 

640 


CHAPTER  37 — DISTRICT  HEATING 


PIPE  CONDUITS 

Conduits  for  steam  pipes  buried  underground  should  be  reasonably 
waterproof,  able  to  withstand  earth  loads  and  to  take  care  of  the  expan- 
sion and  contraction  of  the  piping  without  strain  or  stress  on  the  couplings, 
or  without  affecting  the  insulation  or  conduit.  Expansion  of  the  piping 
must  be  carefully  controlled  by  means  of  anchors  and  expansion  joints 
or  bends  so  that  the  pipes  can  never  come  in  contact  with  the  conduit. 
Anchors  can  be  anchor  fittings  or  U-shaped  steel  straps  which  partially 
encircle  the  pipes  and  are  firmly  bolted  to  a  short  length  of  structural 
steel  set  in  concrete. 


TABLE  1. 


CAPACITY  OF  RETURNS  FOR  UNDERGROUND  DISTRIBUTION  SYSTEMS  IN 
POUNDS  OF  CONDENSATE  PER  HOUR 


SlZEa 

PITCH  OF  PIPE  PER  100  FT. 

OF  PIPE 

IN. 

6" 

i' 

2' 

3' 

5' 

10' 

20' 

1 

448 

998 

1890 

2240 

3490 

5490 

7490 

1M 

1740 

2490 

3990 

4880 

6480 

9480 

13500 

1J4 

2700 

4190 

5740 

7480 

9480 

14500 

20900 

2 

4980 

7380 

10700 

13900 

16900 

24900 

36900 

3 

13900 

22500 

30900 

37400 

50400 

74800 

105000 

4 

30900 

44800 

64800 

79700 

105000 

154000 

229000 

5 

54800 

79800 

120000 

144800 

195000 

294000 

418000 

6 

90000 

138000 

187000 

237000 

312000 

449000 

8 

190000 

277000 

404000 

508000 

660000 

938000 

10 

344000 

498000 

724000 

900000 

1190000 

12 

555000 

798000 

1148000 

1499000 

1990000 





•Size  of  pipe  should  be  increased  if  it  carries  any  steam. 

In  laying  out  conduits  of  this  type  the  following  points  should  be 
borne  in  mind: 

1.  An  expansion  joint  offset  or  bend  should  be  placed  between  each  two  anchors. 

2.  If  the  distance  between  buildings  is  150  ft  or  less  and  the  steam  line  contains  high- 
pressure  steam,  the  line  may  be  anchored  in  the  basement  of  one  building  and  allowed  to 
expand  into  the  basement  of  the  second  building.    If  the  steam  line  contains  low-pressure 
steam  (up  to  4-lb  pressure),  this  method  may  be  used  if  buildings  are  250  ft  or  less  apart. 

3.  If  the  distance  between  buildings  is  between  150  ft  and  300  ft  and  the  steam  line 
contains  high-pressure  steam,  the  lines  should  be  anchored  midway  between  the  buildings 
and  allowed  to  expand  into  the  basements  of  both  buildings.    If  the  steam  line  contains 
low-pressure  steam  this  method  may  be  used  if  buildings  are  between  250  ft  and  600  ft 
apart.    No  manhole  is  required  at  the  anchor,  and  a  blind  pit  is  all  that  is  necessary, 

4.  For  longer  lines,  manholes  must  be  located  according  to  judgment  and  depending 
upon  the  expansion  value  of  the  type  of  expansion  joint  or  bend  that  is  used.    The 
minimum  number  of  manholes  will  be  required  when  an  expansion  bend  or  an  anchor 
with  double  expansion  joint  is  placed  in  each  manhole  and  the  pipes  are  anchored  mid- 
way between  manholes. 

5.  A  proper  hydrostatic  test  should  be  made  on  the  assembled  line  before  the  insula- 
tion and,  the  top  of  the  conduit  are  applied.    The  hydrostatic  pressure  should  be  one- 
and-one-half  times  the  maximum  allowable  pressure  and  it  should  be  held  for  a  period  of 
at  least  two  hours  without  evidence  of  leakage.    In  any  case  the  pressure  should  be  no- 
less  than  100  Ib  per  square  inch. 

The  styles  and  construction  of  conduits  commonly  used  may  be  classi- 
fied as  follows.    Some  of  the  more  common  forms  are  illustrated  in  Fig.  1. 

Wood  Casing:   The  pipe  is  enclosed  in  a  cylindrical  casing  usually  having  a  wall  4  in. 
thick  and  built  of  segments  which  are  bound  together  by  a  wire  wrapped  spirally  around 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

the  casing.  The  casing  is  lined  with  bright  tin  and  coated  with  asphaltum.  The  pipe  is 
supported  on  rollers  carried  in  a  bracket  which  fits  into  the  casing.  The  lengths  of 
casing  are  tightly  fitted  together  with  a  male  and  female  joint.  This  form  of  conduit  is 
illustrated  in  Fig.  1  at  A .  The  casing  rests  on  a  bed  of  crushed  stone  with  tile  drains  laid 
below.  The  tile  drains  are  of  4-in.  field  tile  or  vitrified  sewer  tile,  laid  with  open  joints. 

Filler  Type:  The  pipes  are  supported  on  expansion  rollers  properly  supported  from 
the  conduit  or  independent  masonry  base.  The  pipes  are  protected  by  a  split-tile  conduit, 
and  the  entire  space  between  the  pipes  and  the  tile  is  filled  with  an  insulating  filler.  Thus 
the  pipes  are  nested  and  the  insulation  between  them  and  the  tile  effectively  prevents 
circulation  of  air.  The  conduit  is  placed  on  a  bed  of  gravel  or  crushed  rock  from  4  to  6  in. 
thick,  which  is  extended  upward  so  as  to  come  about  2  in.  above  the  parting  lines  of  the 
tile.  A  tile  underdrain  is  placed  beneath  the  conduit  throughout  the  entire  length  and  is 
connected  to  sewers  or  to  some  other  point  of  free  discharge.  At  B  and  D  in  Fig.  1  are 
shown  two  forms  of  tile  conduit  of  the  filler  type. 

Circular  Tile  or  Cast-Iron  Conduit:  The  pipes  are  carried  on  expansion  rollers  sup- 
ported on  a  frame  which  rests  entirely  on  the  side  shoulders  of  the  base  drain  foundation. 


FIG,  1.    CONSTRUCTION  DETAILS  OF  CONDUITS  COMMONLY  USED 


A  he  pipes  are  protected  by  a  sectional  tile  conduit,  scored  for  splitting,  or  a  cast-iron 
conduit,  both  being  of  the  bell  and  spigot  type.  The  conduit  has  a  longitudinal  side  joint 
for  cementing,  after  the  upper  half  of  conduit  is  in  place,  so  shaped  that  the  cement  is 
keyed  in  place  while  locking  the  top  and  bottom  half  of  the  conduit  together  with  a 
water-tight  vertical  side  joint.  The  cast-iron  conduit  has  special  side  locking  clamps  in 
addition  to  the  vertical  side  joint.  The  entire  space  between  the  conduit  and  the  pipes  is 
filled  with  a  water-proofed  asbestos  insulation.  The  conduit  is  supported  on  the  base 
drain  foundation,  each  section  resting  on  two  sections  of  the  base  drain,  thus  inter- 
locking, The  base  drain  is  so  shaped  that  it  provides  a  cradle  for  the  conduit*  resting 
solidly  on  the  trench  bottom  and  providing  adequate  drainage  area  immediately  under  the 
conduit.  The  underdrain  is  connected  to  sewers  or  some  other  point  of  free  discharge. 
For  tile  conduit  the  base  drain  is  vitrified  salt  glazed  tile  and  for  cast-iron  conduit  it  is 
either  extra  heavy  tile  or  cast-iron.  A  free  internal  drainage  area  is  also  provided  to  carry 
away  any  water  that  may  collect  on  the  inside  of  the  conduit  from  a  leaky  pipe  or  joint  in 
the  conduit.  Broken  stone  is  filled  in  around  the  base  drain  and  up  to  the  vertical  side 
joint.  The  broken  stone  is  covered  with  an  asphalted  filter  cloth  to  prevent  sand 
from  sifting  through  the  broken  stone  and  clogging  the  drainage  area  of  the  base  drain. 
The  tile  conduit  is  made  in  2- ft  lengths  and  the  cast-iron  conduit  in  4-ft  lengths,  cast  in 

642 


CHAPTER  37 — DISTRICT  HEATING 


separate  top  and  bottom  halves.  Special  reinforcing  ribs  give  the  cast-iron  conduit  ample 
strength  with  minimum  weight. 

Insulated  Tile  Type:  The  insulating  material,  diatomaceous  earth,  is  molded  to  the 
inside  of  the  sectional  tile  conduit.  The  space  between  the  pipes  and  the  insulating  con- 
duit lining  may  also  be  filled  with  insulation.  The  pipes  are  carried  on  expansion  rollers 
supported  on  a  frame  which  rests  on  the  side  shoulders  of  the  base  drain  foundation. 
This  type  of  conduit  has  the  same  mechanical  features  as  those  described  under  the 
heading  Circular  Tile  or  Cast-Iron  Conduit. 

Sectional  Insulation  Type  (Tile  or  Cast-iron):  Each  pipe  is  insulated  in  the  usual  way 
with  any  desired  type  of  sectional  pipe  insulation  over  which  is  placed  a  standard  water- 
proof jacket  with  cemented  joints.  The  pipes  are  enclosed  in  a  sectional  tile  or  cast-iron 
conduit  as  described  under  the  heading  Circular  Tile  or  Cast-Iron  Conduits. 

Sectional  Insulation  Type  (Tile  or  Concrete  Trench) :  A  type  of  construction  frequently 
used  in  city  streets,  where  service  connections  are  required  at  frequent  intervals,  the 
pipes  are  insulated  as  described  in  the  preceding  paragraph,  and  are  enclosed^  in  a  box 
or  trench  made  either  entirely  of  concrete,  or  with  concrete  bottom  and  specially  con- 
structed tile  sides  and  tops.  The  pipes  are  supported  on  roller  frames  secured  in  the 
concrete.  At  C  and  E,  Fig.  1,  are  shown  two  tile  conduits  using  sectional  insulation.  In 
these  particular  designs  the  space  surrounding  the  pipe  is  filled  partially  or  wholly  with  a 
loose  insulating  material.  The  use  of  loose  material  in  addition  to  the  sectional  insula- 
tion is,  of  course,  optional  and  is  only  justifiable  where  high  pressure  steam  is  used.  The 
conduit  shown  at  F  is  of  a  similar  type  and  has  the  advantage  of  being  made  entirely  of 
concrete  and  other  common  materials. 

Sectional  Insulation  Type  (Bituminized  Fibre  Conduit) :  Each  pipe  is  individually 
insulated  and  encased  in  a  bituminized  fibre  conduit.  The  insulating  material  is  85 
per  cent  carbonate  of  magnesia  sectional  pipe  covering,  applied  in  the  usual  manner  as 
on  overhead  pipes,  except  that  bands  are  omitted.  After  every  fifth  section  of  magnesia 
covering  there  is  applied  a  short,  hollow  section  of  very  hard  asbestos  material  in  the 
bottom  portion  of  which  rests  a  grooved-iron  plate  carrying  ball-bearings ^  upon  which 
the  pipe  rides  when  expanding  or  contracting.  This  short  expansion  section  is  of  the 
same  outside  diameter  as  the  adjacent  85  per  cent  magnesia  covering.  Over  the  pipe 
covering  and  expansion  device  there  are  placed  two  layers  of  bituminized ( fibre  conduit 
with  all  joints  staggered,  and  the  surface  of  each  conduit  is  finished  with  liquid  cement. 
Conduits  are  placed  on  a  bed  of  crushed  rock  or  gravel,  approximately  6  in.  deep,  and 
this  is  extended  upward  to  about  the  center  line  of  the  conduit  when  trench  is  backfilled. 
Underdrains  leading  to  points  of  free  discharge  are  placed  in  the  gravel  or  crushed 
rock  beds. 

Special  Water-Tight  Designs:  It  is  occasionally  necessary  to  install  pipes  in  a  very  wet 
ground,  which  calls  for  special  construction.  The  ordinary  tile  or  concrete  conduit  is  not 
absolutely  water  tight  even  when  laid  with  the  utmost  care.  The  conduit  shown  at  G, 
Fig,  1,  is  of  cast-iron  with  lead-calked  joints  and  is  water  tight  if  properly  laid.  It  is 
obviously  expensive  and  is  justified  only  in  exceptional  cases.  A  reasonably  satisfactory 
construction  in  wet  ground  is  the  concrete  or  tile  conduit  with  a  waterproof  jacket 
enclosing  the  pipe  and  its  insulation,  and  with  the  interior  of  the  conduit  carefully 
drained  to  a  manhole  or  sump  having  an  automatic  pump.  It  is  useless  to  install  external 
drain  tile  when  the  conduit  is  actually  submerged.  • 

PIPE  TUNNELS 

Where  steam  heating  lines  are  installed  in  tunnels  large  enough  to 
provide  walking  space,  the  pipes  are  supported  by  means  of  hangers  or 
roller  frames  on  brackets  or  frame  racks  at  the  side  or  sides  of  the  tunnel. 
The  pipes  are  insulated  with  sectional  pipe  insulation  over  which  is 
placed  a  sewed-on,  painted  canvas  jacket  or  a  jacket  of  asphalt-saturated 
asbestos  water-proofing  felt.  The  tunnel  itself  is  usually  built  of  concrete 
or  brick  and  water-proofed  on  the  outside  with  membrane  water-proofing. 

On  account  of  their  relatively  high  first  cost  as  compared  with  smaller 
conduits,  walking  tunnels  are  sometimes  not  installed  where  provision  for 
the  heating  lines  is  the  only  consideration,  but  only  where  they  are  required 

643 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

to  accommodate  miscellaneous  other  services  or  provide  underground 
passage  between  buildings. 

SERVICE  CONNECTIONS 

Most  district  heating  companies  enforce  certain  regulations  regarding 
the  consumer's  installation,  partly  to  safeguard  their  own  interests  but 
principally  to  insure  satisfactory  and  economical  service  to  the  consumer. 


Heating  mam 
Unions 


Pipe  to  connect  into  steam  mam  not  less 
than  10  feet  from  reducing  valve,  i(  possible 


FIG.  2.    CONNECTIONS  FOR  REDUCING  VALVES  OF  SIZE  LESS  THAN  4  INCHES 

There  are  certain  fundamental  principles  that  should  be  followed  in  the 
design  of  a  building  heating  system  which  is  to  be  supplied  from  street 
mains.  Although  some  of  these  apply  to  any  building,  they  have  been 
demonstrated  to  be  especially  important  when  steam  is  purchased. 


Bypass  valve' 


Heating  mam 


Service  valve 


Balance  pipe 

Pipe  to  connect  into  steam  mam  not  less 
than  10  feet  from  reducing  valve,  if  possible 


FIG.  3.    CONNECTIONS  FOR  REDUCING  VALVES  OF  SIZE  4  INCHES  AND 
LARGER,  AND  FOR  EXPANDED  VALVES 


Figs.  2  and  3  show  typical  service  connections  used  for  low  pressure 
steam  service.  As  shown  in  Fig.  2,  no  by-pass  is  used  around  the  reducing 
valve  on  sizes  less  than  4  in*  Fig.  3  illustrates  the  use  of  a  by-pass  around 
reducing  valves  4  in.  and  larger.  This  latter .  construction  permits  the 


644 


CHAPTER  37 — DISTRICT  HEATING 


operation  of  the  line  in  case  of  failure  in  the  reducing  valve.  In  the  smaller 
sizes,  the  reducing  valve  can  be  removed,  a  filler  installed,  and  the  house 
valve  used  to  throttle  the  flow  of  steam. 

Fig.  4  shows  a  typical  installation  used  for  high  pressure  steam  service. 
The  first  reducing  valve,   usually  furnished  by  the  utility  company, 


Pressure  reducing  valve 


At  least  12  feet  of  pipe 


Customer's  work 
starts  here 


Note.-  All  valves,  fittings,  and  traps  up  to 
and  including  customer's  control 

valve  to  be  at  least  equal  to 
American  Standard  175  Ib  S.  S  P. 
Pipe  to  be  standard  weight 


Continuous-flow  type 
float  trap 


FIG.  4.    STEAM  SUPPLY  CONNECTION  WHEN  USING  CONDENSATION  METER 

effects  the  initial  pressure  reduction.  The  second  reducing  valve,  usually 
furnished  by  the  customer,  reduces  the  steam  pressure  to  that  required. 

1 .  Provision  should  be  made  for  conveniently  shutting  off  the  steam  supply 
at  night  and  at  other  times  when  heat  is  not  needed. 

It  has  been  thoroughly  demonstrated  that  a  considerable  amount  of 
heat  can  be  saved  by  shutting  off  steam  at  night.  Although  there  is,  in 


Return  main 


Condensation  meter 
and  manifold  castmi 


Vent 


[V-  Preheated  water 
to  heater 


-   -Carry  full  size  to  sewer 
• — Gas  seal 

FIG.  5.    RETURN  PIPING  FOR  CONDENSATION  METER 

some  cases,  an  increased  consumption  of  heat  when  steam  is  again  turned 
on  in  the  morning,  there  is  a  large  net  saving  which  may  be  explained  by 
the  fact  that  the  lower  inside  temperature  maintained  during  the  night 
obviously  results  in  lower  heat  loss  from  the  building,  and  less  heat  need 
therefore  be  supplied. 
Steam  can  be  entirely  shut  off  at  night  in  most  buildings  even  in  very 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


CHAPTER  37 — DISTRICT  HEATING 


cold  weather  without  endangering  plumbing.  It  is  necessary,  however,  to 
have  an  ample  amount  of  heating  surface  so  that  the  building  can  be 
quickly  warmed  in  the  morning.  Where  the  hours  of  occupancy  differ  in 
various  parts  of  the  building,  it  is  good  practice  to  install  separate  supply 
pipes  to  the  different  parts.  For  example,  in  an  office  building  with 
stores  or  restaurants  on  the  first  floor  which  are  open  in  the  evening,  a 
separate  main  supplying  the  first  floor  will  permit  the  steam  to  be  shut  off 
from  the  remainder  of  the  building  in  the  late  afternoon.  The  division  of 
the  building  into  zones  each  with  a  separately  controlled  heat  supply  is 
sometimes  desirable,  as  it  permits  the  heat  to  be  adjusted  according  to 
variations  in  sunshine  and  wind. 

#.    Residual  heat  in  the  condensate  should  be  salvaged. 

This  heat  may  be  salvaged  by  means  of  a  cooling  radiator,  or  as  is  more 
frequently  done,  by  a  water  heating  economizer  (see  Fig.  5)  which  pre- 
heats the  hot  water  supply  to  the  building.  Fig.  6  shows  a  typical  steam 
service  installation  for  high  pressure  steam,  complete  for  steam  flow 
metering,  water  heating,  preheating,  automatic  heating  control,  and  for 
using  steam  for  other  purposes. 

The  condensation  from  the  heating  system,  after  leaving  the  trap, 
passes  through  the  preheater  on  its  way  to  the  meter.  The  supply  to  the 
hot  water  heater  passes  through  the  preheater,  absorbing  heat  from  the 
condensation.  If  the  hot  water  system  in  the  building  is  of  the  recircu- 
lating  type,  the  recirculating  connection  should  be  tied  in  between  the 
preheater  and  the  water  heater  proper,  not  at  the  preheater  inlet,  because 
the  recirculated  hot  water  is  itself  at  a  high  temperature.  The  number  of 
square  feet  of  heating  surface  in  the  preheater  should  be  approximately 
equal  to  one  per  cent  of  the  equivalent  square  feet  of  heating  surface  in  the 
building. 

Because  of  the  lack  of  coincidence  between  the  heating  system  load  and 
the  hot  water  demand,  a  greater  amount  of  heat  can  be  extracted  from  the 
condensation  if  storage  capacity  is  provided  for  the  preheated  water. 
Frequently  a  type  of  preheater  is  used  in  which  the  coils  are  submerged 
in  a  storage  tank. 

8.  Heat  supply  should  be  graduated  according  to  variations  in  the  outside 
temperature. 

This  may  be  done  in  several  ways,  as  by  the  use  of  thermostats  of 
various  types  or  by  orifice  systems.  Another  method  which  is  very  simple 
is  the  use  of  an  ordinary  vacuum  return  line  system  in  which  the  pressure 
in  the  radiators  is  varied  between  a  high  vacuum  and  a  few  pounds  pres- 
sure, thus  producing  some  control  over  the  heat  output.  One  form  of  con- 
trol which  appears  to  be  well  suited  for  controlling  district  steam  service 
to  a  building  is  the  weather  compensating  thermostat.  It  regulates  the 
steam  supply  automatically  according  to  the  outdoor  temperature,  and 
gives  frequent  short  intervals  of  intermittent  steam  supply,  and  at  the 
same  time  insures  delivery  of  steam  to  all  the  radiators. 

Another  form  of  regulation,  known  as  the  time-limit  control,  is  sometimes 
employed  for  regulating  the  steam  supply  from  the  central  station  main  to 
the  building.  Such  a  control  provides  an  intermittent  supply  of  steam  to 
the  radiation  either  throughout  the  24  hours  of  the  day  or  during  the  day- 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


time  hours  only.  The  setting  of  a  switch  may  provide  no  service,  con- 
tinuous service,  or  periodic  service.  For  the  latter,  by  means  of  several 
intermittent  settings,  steam  will  be  supplied  during  each  period  in  in- 
crements of  a  certain  number  of  minutes  for  each  successive  setting  of  the 
switch,  steam  being  shut  off  during  the  balance  of  the  period.  These 
settings  afford  from  15  to  80  per  cent  of  the  maximum  heating  effect 
required  on  days  of  zero  temperature.  A  night  switch  with  a  variety  of 
settings  may  be  adjusted  so  as  to  maintain  throughout  the  night  the 
intermittent  supply  called  for  by  the  day  switch  setting,  or  may  be  set  to 
interrupt  the  operation  of  the  day  switch  and  entirely  cut  off  the  supply 
of  steam  to  the  radiation  at  night  during  certain  hours  which  are  selected 
by  the  operating  engineer. 

FLUID  METERS 

No  one  thing  has  contributed  more  to  the  advancement  of  district 
heating  than  the  perfection  of  fluid  meters,  which  may  be  classified  as 
follows : 

1.  Positive  Meters:  The  fluid  passes  in  successive  isolated  quantities — either  weights 
or  volumes.  These  quantities  are  separated  from  the  stream  and  isolated  by  alternately 
filling  and  emptying  containers  of  known  capacity. 

2.  Differential  Meters:  The  fluid  does  not  pass  in  isolated  separately-counted  quan- 
tities but  in  a  continuous  stream  which  may  flow  through  the  line  without  actuating 
the  primary  device  of  the  meter.   In  the  differential  meter,  the  quantity  of  flow  is  not 
determined  by  simple  counting,  as  with  the  positive  meter,  but  is  determined  from  the 
action  of  the  steam  on  the  primary  element. 

Additional  subdivisions  of  these  two  general  classifications  can  be  made 
as  follows : 


Fluid 
Meters 


Positive  -  quantity 

Weighing 
Volumetric 

f     Weighers 
\     Tilting  trap 

/     Rotary 
\     Bellows 

Quantity  -  Current  -  Turbine 

Ferential     < 

Rate  of 
flow 

Head 
(Kinetic) 

Area 
(Geometric) 

IVenturi 
Flow  nozzle 
Orifice 
Pitot  tube 

(    Orifice  and  plug 
\    Cylinder  and  piston 

Head  area 
(Weir) 

f     V-notch 
\     Special  notch 

In  selecting  a  meter  for  a  particular  installation,  the  number  of  different 
makes  and  types  of  meters  suitable  for  the  job  is  usually  limited  by 
one  or  more  of  the  following  considerations: 

1.  Its  use  in  a  new  or  an  old  installation. 

2.  Method  to  be  used  in  charging  for  the  service. 

3.  Location  of  the  meter. 

4.  Large  or  small  quantity  to  be  measured. 

5.  Temporary  or  permanent  installation. 

648 


CHAPTER  37 — DISTRICT  HEATING 


Pressure  reducing  valve 


and  zap 


Vents  and  loops  unnecessary 

where  meter  is  5  feet  or 

more  below  pipe 


Note.-  All  valves,  fittings,  and  traps  up  to 
and  including  customer's  control 

valve  to  be  at  least  equal  to 
American  Standard  175  Ib  S.  S,  P 
Pipe  to  be  standard  weight 


FIG.  7.    ORIFICE  METER  STEAM  SUPPLY  CONNECTION 


6.  Cleanliness  of  the  fluid  to  be  measured. 

7.  Temperature  of  the  fluid  to  be  measured. 

8.  Accuracy  expected. 

9.  Nature  of  flow:  turbulent,  pulsating,  or  steady. 

10.  Cost. 

(a)  Purchase  price. 
(&)   Installation  cost. 

(c)  Calibration  cost. 

(d)  Maintenance  cost. 

11.  Servicing  facilities  of  the  manufacturer. 

12.  Pressure  at  which  fluid  is  to  be  metered. 

13.  Type  of  record  desired  as  to  indicating,  recording  or  totalizing. 

14.  Stocking  of  repair  parts. 

15.  Use  of  open  jets  where  steam  is  to  be  metered. 

16.  Metering  to  be  done  by  one  meter  or  by  a  combination  of  meters. 

17.  Use  as  a  check  meter. 

18.  Its  facilities  for  determining  or  recording  information  other  than  flow. 

Condensation  Meters 

The  majority  of  the  meters  used  by  district  heating  companies  in  the 
sale  of  steam  to  their  customers  are  of  the  condensation  or  flow  types. 

The  condensation  meter  is  a  popular  type  for  use  on  small  and  medium 
sized  installations,  wnere  a^  °f tne  condensate  can  be  brought  to  a  com- 
mon point  for  metering  purposes.  Its  simplicity  of  design,  ease  in  testing, 
accuracy  at  all  loads,  low  cost,  and  adaptability  to  low  pressure  distri- 
bution has  made  it  standard  equipment  with  many  heating  companies. 

Two  types  of  condensation  meters  are  in  general  use :  the  tilting  bucket 
meter  and  the  revolving  drum  or  rotor  meter  of  which  there  are  several 
makes  on  the  market.  Condensation  meters  should  not  be  operated  under 

649 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

pressure;  they  are  made  for  either  gravity  or  vacuum  installation.  Con- 
tinuous flow  traps  are  necessary  ahead  of  the  meter  if  a  vented  receiver  is 
not  used.  Where  bucket  traps  are  used,  a  vented  receiver  before  the 
meter  is  essential.  If  desirable  a  receiver  may  be  used  with  a  continuous 
flow  trap,  but  this  is  not  necessary. 

Steam  flow  meters  are  available  in  many  types  and  combinations,  as 
indicated  in  the  subdivision  covering  fluid  meters  on  page  648. 

The  orifice  and  plug  meter  is  one  in  which  the  steam  flow  varies  directly 
as  the  area  of  the  orifice.  The  vertical  lift  of  the  plug,  which  is  proportional 
to  the  flow,  is  transmitted  by  means  of  a  lever  to  an  indicator  and  to  a 


CONSTANT  PLOW  TRAP, 


VENT 


GRAVITY  TYPE 

CONDENSATION 

METER-? 


GRAVITY  DISCHARGE 


UJ 


FIG.  8.  GRAVITY  INSTALLATION  FOR  CONDENSATION  METER 
USING  VENTED  RECEIVERS 

pencil  arm  which  records  the  flow  on  a  strip  chart.  The  total  flow  over  a 
given  period  is  obtained  by  measuring  the  area  by  using  a  plani meter 
on  the  chart  and  applying  the  meter  constant. 

Fig.  7  shows  a  typical  orifice  type  meter  connection  and  indicates 
typical  requirements  in  the  installation  of  this  type  of  meter.  Fig.  8 
illustrates  a  gravity  installation  using  a  vented  receiver  ahead  of  the 
meter,  while  Fig.  9  shows  a  vacuum  installation  without  a  master  trap. 

Flow  meters  using  an  orifice,  Venturi  tube,  flow  nozzle,  or  Pi  tot  tube 
as  the  primary  device  are  made  by  a  number  of  manufacturers  and  can 
be  obtained  in  either  the  mechanically  or  electrically  operated  type.  The 
electric  flow  meter  makes  it  possible  to  locate  the  instruments  at  some 
distance  from  the  primary  element* 

Flow  meters  employing  the  orifice,  Venturi  tube,  flow  nozzle  or  Pitot 
tube  should  be  so  selected  as  to  keep  the  lower  operating  range  of  the 
load  above  20  per  cent  of  the  capacity  of  the  meter.  This  is  desirable  for 
accuracy  as  the  differential  pressure  at  light  loads  is  too  small  to  properly 
actuate  the  meter.  A  few  general  points  to  be  considered  in  installing  a 
meter  of  this  type  are: 

1.  It  is  desirable  to  place  the  differential  medium  in  a  horizontal  pipe  in  preference 
to  a  vertical  one,  where  either  location  is  available. 

2,  Reservoirs  should  always  be  on  the  same  level  and  installed  in  accordance  with  the 
instructions  of  the  meter  company. 

660 


CHAPTER  37 — DISTRICT  HEATING 


3.  The  meter  body  should  be  placed  at  a  lower  level  than  that  of  the  pressure  differ- 
ential medium.  Special  instructions  are  furnished  where  the  meter  body  is  above. 

4.  Meter  piping  should  be  kept  free  from  leaks. 

5.  Sludge  should  not  be  permitted  to  collect  in  the  meter  body. 

6.  The  meter  body  and  meter  piping  should  be  kept  above  freezing  temperatures. 

7.  It  is  best  not  to  connect  a  meter  body  to  more  than  one  service. 

8.  Special  instructions  are  furnished  for  metering  a  turbulent  or  pulsating  flow. 

STEAM  PER  SQUARE  FOOT  OF  HEATING  SURFACE 

The  following  factors  are  used  in  New  York  City  for  the  different  classes 
of  buildings  listed.    The  factors  are  based  on  maintaining  an  inside  tern- 


VACUUM 
LINE    FROM 
RADIATORS 


BY-PASS 


VACUUM  TYPE 
CONDENSATION 
METER- 


-TO  VACUUM  PUMP 
FIG.  9.  VACUUM  CONDENSATION  METER  INSTALLATION  WITHOUT  MASTER  TRAP 

perature  of  70  F  for  certain  hours,  with  a  minimum  outside  temperature  of 

0  F  and  an  average  of  43  F  for  the  heating  season  of  eight  months  (October 

1  to  June  1).    In  this  group  are  six  types  of  buildings: 

Manufacturing  or  commercial  loft  type  where  steam  is  used  to  heat  the  premises  during 
the  day  hours  to  maintain  65  to  68  F  from  9  a.m.  to  5  p.m.  No  Sunday  or  holiday  use 
and  no  night  use.  Factor :  325  Ib  per  square  foot  of  heating  surface  per  season, 

Office  buildings  using  steam  during  daylight  hours  to  maintain  70  F  from  9  a.m.  to 
6  p.m.  for  approximately  240  days  (heating  season),  No  night  use,  Factor:  400  Ib  per 
square  foot  of  heating  surface  per  season. 

Office  buildings  using  steam  during  day  hours  and  at  night  when  required  to  7,  8  and 
9  p.m.  (customary  where  there  are  stock  brokers  or  banking  offices),  240  days.  Factor: 
500  Ib  per  square  foot  of  heating  surface  per  season. 

Residences  of  the  block  type  (not  detached)  where  high-class  heatir*g  service  is  re- 
quired; somewhat  similar  to  apartment  buildings.  .  Factor:  550  Ib  per  square  foot  of 
heating  surface  per  season, 

651 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

Apartment  houses  where  high-class  heating  service  is  required.  (Steam  off  at  mid- 
night.) Factor:  650  Ib  per  square  foot  of  heating  surface  per  season. 

Hotels  (commercial  type)  where  very  high-class  service  is  required  for  24  hours. 
Factor:  800  Ib  per  square  foot  of  heating  surface  per  season. 

By  assuming  one  square  foot  of  equivalent  heating  surface  for  each 
100  cu  ft  of  space  heated,  which  seems  a  fair  ratio  in  New  York  City,  it  is 
possible  roughly  to  estimate  the  steam  required  per  cubic  foot  of  space, 
information  which  is  often  more  easily  obtained  than  the  square  feet  of 
heating  surface.  Additional  data  on  the  heating  requirements  of  various 
types  of  buildings  in  a  number  of  cities  may  be  found  in  the  Handbook 
of  the  National  District  Heating  Association. 

RATES 

Fundamentally,  district  heating  rates  are  based  upon  the  same  princi- 
ples as  those  recognized  in  the  electric  light  and  power  industry,  the  main 
object  being  a  reasonable  return  on  the  investment.  However,  there  are 
other  requirements  to  be  met ;  the  rate  for  each  class  of  service  should  be 
based  upon  the  cost  to  the  utility  company  of  the  service  supplied  and 
upon  the  value  of  the  service  to  the  consumer,  and  it  must  be  between 
these  two  limits.  The  profit  need  not  be  divided  proportionately  among 
the  rated  groups,  but  should  be  established  from  a  competitive  stand- 
point. District  heating  rates  should  be  designed  to  produce  a  sufficient 
return  on  the  investment  regardless  of  weather  conditions,  although 
existing  rate  schedules  do  not  conform  with  this  principle.  Lastly,  the 
rate  schedule  must  be  reasonably  easy  for  the  intelligent  layman  to 
comprehend. 

Depreciation  should  be  based  on  a  careful  estimate  of  the  life  of  various 
elements  of  the  property.  Appropriations  to  reserves  should  be  made, 
with  generosity  in  good  years  and  with  discretion  in  less  favorable  years. 

Glossary  of  Terms 

Load  Factor.  The  ratio,  in  per  cent,  of  the  average  load  to  the  maxi- 
mum load.  This  is  usually  based  on  a  one  year  period  but  may  be  applied 
to  any  specified  period. 

Demand  Factor.  The  relation  between  the  connected  radiator  surface 
or  required  radiator  surface  and  the  demand  of  the  particular  installation, 
It  varies  from  0.25  to  0.3  Ib  per  hour  per  square  foot  of  surface. 

Diversity  Factor,  The  ratio  of  the  sum  of  the  individual  demands  of  a 
number  of  buildings  to  the  actual  composite  demand  of  the  group. 

Types  of  Rates 

A.  Flat  Rates. 

1.  Radiator  surface  charge.    Obsolescent, 

JB.  Meter  Rates. 

1.  Straight-line. 

2.  Step.    Obsolescent. 

3.  Block. 

(a)  Class  rates. 

C.  Demand  Rates. 

1.  Flat  demand, 

2.  Wright. 

3.  Hopkinsoru 

4.  Doherty  (or  Three  charge). 

652 


CHAPTER  37 — DISTRICT  HEATING 


Straight-Line  Meter  Rate.  The  price  charged  per  unit  is  constant,  and  the  consumer 
pays  in  direct  proportion  to  his  consumption  without  regard  to  the  difference  in  costs  of 
supplying  the  individual  customers. 

Block  Meter  Rate,  The  pounds  of  steam  consumed  by  a  customer  are  divided  into 
blocks  of  M  pounds  each,  and  lower  rates  are  charged  for  each  successive  block  consumed. 
This  type  of  charge  predominates  in  steam  heating  rate  schedules  for  it  has  the  ad- 
vantage of  proportioning  the  bill  according  to  the  consumption  and  the  cost  of  service. 
It  has  the  disadvantage  of  not  discriminating  between  customers  having  a  high  load 
factor  (relatively  low  demand)  and  those  haying  a  low  load  factor  (relatively  high 
demand).  The  utility  company  must  maintain  sufficient  capacity  to  serve  the  high 
demand  customers  and  the  cost  of  the  increased  plant  investment  is  divided  equally 
among  the  users,  so  the  high  demand  customers  are  benefited  at  the  expense  of  the 
others. 

Demand  Rates.  These  refer  to  any  method  of  charge  based  on  a  measured  maximum 
load  during  a  specified  period  of  time. 

The  flat  demand  rate  is  usually  expressed  in  dollars  per  M  Ib  of  demand  per 
month  or  per  annum.  It  is  based  on  the  size  of  a  customer's  installation,  and  is 
seldom  used  except  where  a  flow  meter  is  not  practicable. 

The  Wright  demand  rate  is  similar  in  calculation  to  the  block  rate  except  that  it  is 
expressed  in  terms  of  hours'  use  of  the  maximum  demand.  It  is  seldom  used  but 
forms  the  basis  for  other  forms  of  rates. 

The  Hopkinson  demand  rate  is  divided  into  two  elements: 

(a)  A  charge  based  upon  the  demand,  either  estimated  or  measured; 

(&)  A  charge  based  upon  the  amount  of  steam  consumed. 

This  rate  may  be  modified  by  dividing  the  quantities  of  steam  demanded  and 
consumed  into  blocks  charged  for  at  different  rates. 

Demand  rates  are  comparatively  new  and  are  not  yet  widely  used;  though  they  are 
equitable  and  competitive  they  are  difficult  for  the  average  layman  to  understand. 
They  are  of  benefit  to  utility  companies  and  to  consumers  because  the  investment  and 
operating  costs  can  be  divided  to  suit  the  particular  circumstances  into  demand,  cus- 
tomer, and  consumption  groups  through  the  use  of  some  modification  of  the  Hopkinson 
rate, 

Fuel  Price  Surcharge,  It  is  usually  desirable  to  establish  a  rate  upon  a  specified  basic 
cost  of  fuel  to  the  utility  company.  Where  there  are  wide  variations  in  the  price  of  fuel, 
it  is  also  desirable  to  add  a  definite  charge  per  M  Ib  of  steam  sold  for  each  increment  of 
increase  in  the  price  of  fuel.  This  surcharge  automatically  compensates  for  the  variations 
without  necessitating  frequent  changing  of  the  whole  rate  structure, 

REFERENCES 

Pipe  Line  Design  for  Central  Station  Healing,  by  B.  T.  Gifford  (A.S.H.V.E,  TRANSACTIONS,  Vol.  17, 1911). 

Engineering  and  Cost  Data  Relative  to  the  Installation  of  Steam  Distributing  Systems  in  a  Large  City,  by 
F.  H.  Valentine  (A.S.H.V.E,  TRANSACTIONS,  Vol.  22,  1916). 

Transmission  of  Steam  in  a  Central  Heating  System,  by  J.  H.  Walker  (A.S.H.V.E.  TRANSACTIONS, 
Vol.  23,  1917). 

Efficiency  of  Underground  Conduit,  by  G.  B.  Nichols  (A.S.H.V.E.  TRANSACTIONS,  Vol.  23,  1917). 

Economical  Utilization  of  Heat  from  Central  Plants,  by  N,  W.  Calvert  and  J,  E.  Seiter  (A.S.H.V.E. 
TRANSACTIONS,  Vol.  30,  1924). 

Standard  Connections  for  Condensation  Meiers,  (N.D.H.A.  Proceedings,  Vol.  XII,  pp.  63-76). 

Installation  and  Maintenance  of  Steam  Meters,  (N.D.H.A.  Proceedings,  Vol.  XIII,  pp.  177-183), 

Inaccuracy  in  Flow  Meter  Calculations,  (N.D.B.A.  Proceedings,  Vol.  XIII,  pp.  183-193), 

Testing  of  Steam  Meters,  (N.D.H.A.  Proceedings,  Vol.  XIV,  pp.  272-276). 

Meter  Accuracy  Guarantees,  (N.D.H.A.  Proceedings   Vol.  XIV,  pp.  276-277), 

Effect  of  Pulsations  on  the  Flow  of  Cases,  (N.D.H.A,  Proceedings,  Vol.  XIV,  pp.  277-281). 

Meter  Connections,  (N.D.H.A.  Proceedings,  Vol.  XX,  pp,  126-143). 

Layout  for  Testing  Meters,  (N.D,H,A.  Proceedings,  Vol.  XX,  pp.  391-392), 

Characteristic  Meter  Calibration  Curves  (N.D.ff.A-  Proceedings,  Vol.  XX,  pp.  444-453), 

Rates  (N.D.H.A.  Handbook,  1932,  Chapter  10). 


653 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


PROBLEMS  IN  PRACTICE 

1  •  What  is  the  common  method  of  determining  the  size  of  mains  in  a  dis- 
tribution system? 

On  the  basis  of  pressure  drop:  The  initial  pressure  and  the  minimum  permissible 
terminal  pressure  are  specified,  and  the  pipe  sizes  are  so  chosen  that  the  maximum 
estimated  amount  of  steam  may  be  transmitted  without  exceeding  this  pressure  dif- 
ference. The  steam's  velocity  is  disregarded  and  it  may  reach  a  magnitude  in  excess  of 
35,000  fpm  which  is  not  considered  high. 

2  •  a.  What  are  the  advantages  and  disadvantages  of  a  low  pressure  distribu- 
tion system? 

b.  High  pressure? 

a.  The  advantages  of  a  low  pressure  distribution  system  include: 

1.  Smaller  heat  loss  from  the  pipes. 

2.  Less  trouble  with  traps  and  valves. 

3.  Simpler  problems  with  pressure  reducing  equipment  at  the  buildings. 

4.  No  danger  to  building  heating  equipment  from  high  pressure  through  failure  of  the 
reducing  valves. 

The  disadvantages  of  a  low  pressure  system  are: 

1.  Larger  pipe  sizes. 

2,  Decreased  field  of  usefulness  owing  to  small  pressure  range. 

b.  The  advantages  of  a  high  pressure  system  are: 

1.  Smaller  pipe  sizes. 

2,  Greater  adaptability  of  the  steam  to  various  uses  other  than  building  heating. 
The  disadvantages  of  a  high  pressure  system  are : 

1.  Large  heat  loss  from  the  pipes. 

2.  The  high  pressure  traps  and  valves  required  often  give  more  trouble  than  low 
pressure  traps  and  valves  do. 

3.  Extra  heavy  fittings  are  required. 

4.  Usually  two  reducing  valves  or  some  form  of  emergency  relief  is  necessary  to 
protect  the  building  piping  system. 

3  •  Determine  the  size  of  pipe  from  the  following  data  using  Unwin's  formula: 
Length  of  pipe,  600  ft. 

Steam  to  be  carried,  90,000  Ib  per  hour,  dry  saturated. 
Initial  pressure,  100  Ib  per  square  inch,  gage. 
Final  pressure,  40  Ib  per  square  inch,  gage. 
Using  the  formula: 

0,0001306  W*L  (  1  4-  ~> 

p  «    _.^____i „£, 

yd* 

The  pressure  drop  P  »  100  —  40  »  60  Ib  per  square  inch. 

90  000 
The  weight  of  steam  per  minute  W  =*  — ~— -  **  1500, 

The  length  of  pipe  in  feet  L  »  600. 

The  average  density  of  steam  y  in  pounds  per  cubic  foot,  taken  fyrc-ni  ^teenan's  Tf^ble: 

At  100-lb  gage,  y  -  0.2578 

At   40-lb  gage,  y  »  0.1285 

Average,  y  «  0.1932 

The  diameter  of  the  pipe  in  inches  =»  d. 

654 


CHAPTER  37 — DISTRICT  HEATING 


Substituting  the  values  in  the  formula: 

0.0001306  X  15002  X  600  (  1  +  ^ 

fi0  _  V  & 

°U  0.1932  X  # 

d  «  7.35  in. 
Therefore,  an  8-in.  pipe  should  be  used. 

4  •  What  points  should  be  borne  in  mind  when  laying  out  an  underground 
steam  conduit? 

The  conduit  should  be  reasonably  waterproof,  able  to  withstand  earth  loads  and  to  take 
care  of  the  expansion  and  contraction  of  the  piping  without  strain  or  stress  on  the 
couplings,  or  without  affecting  the  insulation  or  the  conduit.  Expansion  of  the  piping 
must  be  carefully  controlled  by  means  of  anchors  and  expansion  joints  or  bends  so  that 
the  pipes  can  never  come  in  contact  with  the  conduit. 

5  •  What  is  considered  the  proper  pressure  for  a  hydrostatic  test  before  com- 
pleting the  conduit? 

In  the  case  of  any  underground  piping  which  is  to  be  ^buried  or  otherwise  made  inacces- 
sible, the  assembled  lines  shall  first  be  tested  hydrostatically  at  a  pressure  of  one  and  one- 
half  times  the  maximum  allowable  service  pressure  and  held  for  a  period  of  at  least  two 
hours  without  evidence  of  leakage.  In  any  case  the  hydrostatic  pressure  should  not  be 
less  than  100  Ib  per  square  inch. 

6  •  What  factors  should  be  considered  before  determining  the  route  of  a  steam 
line? 

1.  The  line  should  be  so  located  that  it  will  bring  in  the  greatest  revenue  (or  supply  the 
most  steam)  with  the  least  cost. 

2.  The  ultimate  length  and  size  of  services  and  branches  necessary  with  each  possible 
location  should  be  estimated,  for  mains  should  be  run  near  to  the  big  loads. 

3.  The  location  of  the  boiler  room  or  piping  center  of  present  and  future  buildings  to  be 
served  should  be  considered. 

4.  Where  possible,  make  the  lines  straight  between  manholes. 

5.  Avoid  such  obstructions  as  other  lines,  sewers,  ducts,  curb  drains,  manholes,  valve 
boxes,  catch  basins,  fire  hydrants,  and  poles;  especially  avoid  electric  ducts  and  water 
lines. 

6.  Avoid  locating  lines  near  where  pile  driving  and  foundation  construction  for  new 
buildings  will  take  place. 

7.  Consider  construction  difficulties  such  as  traffic,  hard  rock,  and  wet  earth,  which 
increase  time  and  labor, 

8.  Consider  the  economies  of  using  available  sidewalk  vaults  of  buildings.    Weigh  the 
advantage  of  less  excavation  against  the  cost  of  obstruction  removal, 

9.  Consider  all  operating  difficulties, 

10.  Consider  the  difficulties  of  negotiating  agreements  for  lines  on  private  property 
where  public  and  private  rights-of-way  are  available. 

11.  Consider  the  effect  of  proposed  municipal  and  other  improvements. 

12.  Consider  municipal  regulations. 

7.  •  State  the  advantages  and  disadvantages  of  tunnels  over  conduits. 

The  advantages  of  pipe  tunnels  over  conduits  are: 

1.  Accommodation  for  miscellaneous  services  other  than  steam. 

2.  Provision  of  an  underground  passage  between  buildings. 

3.  Easy  installation  of  additional  pipes  and  easy  replacement  of  existing  pipes  with 
larger  sizes. 

4.  Easy  inspection  and  maintenance  of  pipes. 

655 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

The  disadvantages  of  pipe  tunnels  over  conduits  are: 

1.  Higher  first  cost. 

2.  Higher  maintenance  cost  in  general. 

8  •  Is  the  steam  consumption  less  in  a  building  that  shuts  off  its  steam  at 
night  than  in  one  that  does  not?    Why? 

It  has  been  thoroughly  demonstrated  that  the  steam  consumption  is  less  in  a  building 
where  the  steam  is  shut  off  at  night.  Although  therejs,  in  some  cases,  an  increased  con- 
sumption of  heat  when  steam  is  again  turned  on  in  the  morning,  there  is  a  large  net 
saving  which  may  be  explained  by  the  fact  that  the  lower  inside  temperature  maintained 
during  the  night  obviously  results  in  lower  heat  loss  from  the  building,  and  less  heat  need 
therefore  be  supplied. 

9  •  What  are  the  common  methods  for  salvaging  heat  in  condensate? 

The  most  common  methods  are: 

1.  The  use  of  a  water  heating  economizer  for  preheating  the  hot  water  supply  to  the 
building. 

2.  The  use  of  a  cooling  radiator. 

10  •  What  are  the  common  means  used  to  graduate  the  heat  supply  according 
to  variations  in  outside  temperature? 

a.  A  weather  compensating  thermostat  regulates  the  steam   supply  automatically 
according  to  the  outdoor  temperature,  and  gives  frequent  short  intervals  of  inter- 
mittent steam  supply;  at  the  same  time  it  insures  delivery  of  steam  to  all  the  radiators. 

b.  Another  method  which  is  very  simple  is  the  use  of  an  ordinary  vacuum  return  line 
system  in  which  the  pressure  in  the  radiators  is  varied  between  a  high  vacuum  and  a 
few  pounds  to  produce  some  control  over  the  heat  output. 

c.  The  use  of  an  orifice  system  graduates  heat  supply. 

d.  The  time-limit  control  which  may  be  set  to  provide  no  service,  continuous  service,  or 
periodic  service,  is  also  used.    For  periodic  service,  steam  may  be  supplied  during 
each  period  in  increments  of  a  certain  number  of  minutes  for  each  successive  setting 
of  the  switch,  steam  being  shut  off  during  the  balance  of  the  period.    This  type  of 
service  is  provided  by  several  intermittent  settings.    A  night  switch  will  maintain 
the  intermittent  day  setting,  or  interrupt  the  day  operation  and  cut  off  the  supply  of 
steam  at  night  during  any  desired  hours. 


656 


Chapter  38 

RADIANT  HEATING 

Physical    and    Physiological    Considerations,    British    Equivalent 
Temperature,   Control  of  Heat  Losses,  Methods  of  Application, 
Principles  of  Calculation,  Mean  Radiant  Temperature,  Measure- 
ment of  Radiant  Heating 

HEATING  for  comfort  is  generally  understood  to  mean  that  heat 
must  be  supplied  to  control  the  rate  of  heat  loss  from  the  human 
body  so  that  the  physiological  reactions  are  conducive  to  a  feeling  of 
comfort  in  the  individual.  While  in  convection  heating,  as  described  in 
Chapter  30,  heat  is  transferred  from  a  heating  unit  to  the  air  and  thence 
to  the  occupant,  the  primary  object  of  radiant  heating  is  to  warm  the 
occupant  directly  without  heating  the  air  to  any  extent.  Thus,  the 
difference  between  convection  heating  and  radiant  heating  is  partly 
physical  and  partly  physiological. 

Comfort  requires  that  heat  be  removed  from  the  body  at  the  same  rate 
as  it  is  generated  by  the  oxidation  of  the  foodstuffs  in  the  body  tissues. 
The  normal  rate  of  heat  production  in  a  sedentary  individual  is  about 
400  Btu  per  hour1,  or,  since  the  entire  surface  area  of  an  average  adult  is 
19.5  sq  ft,  about  20.5  Btu  per  square  foot  per  hour.  Conditions  should  be 
such  as  to  remove  heat  at  this  rate  if  the  surface  is  to  be  maintained  at  the 
mean  normal  surface  temperature  of  the  human  body. 

Heat  is  transferred  from  any  warm  dry  body  to  cooler  surroundings 
principally  by  convection  and  by  radiation,  the  approximate  total  rate  of 
heat  loss  being  the  sum  of  the  two.  Where  the  body  surface  is  moist  there 
is  additional  loss  of  heat  through  evaporation  from  both  the  body  surface 
and  the  respiratory  tract., 

The  rate  of  heat  loss  by  convection  depends  upon  the  difference  between 
the  temperature  of  the  body  and  that  of  the  surrounding  air,  and  on  the 
rate  of  air  motion  over  the  body.  The  loss  by  radiation  depends  entirely 
upon  the  difference  between  the  temperature  of  the  body  and  the  mean 
surface  temperature  of  the  surrounding  walls  and  objects.  This  latter 
temperature  is  called  the  mean  radiant  temperature  (MRT).  Because 
these  two  types  of  heat  loss  a'ct  in  a  supplementary  manner  toward  each 
other,  a  required  rate  of  heat  loss  can  be  secured  by  having  a  relatively 
low  air  temperature  and  a  relatively  high  MRT,  or  vice  versa.  Thus,  if  the 
air  is  reduced  from  a  given  temperature  to  a  lower  temperature,  the 
amount  of  heat  lost  from  the  body  by  convection  is  increased,  and  this 


lHeat  and  Moisture  Losses  from  the  Human  Body  and  Their  Relation  to  Air  Conditioning  Problems, 
by  F.  C.  Houghten,  W.  W.  Teague,  W.  E.  Miller,  and  W.  P.  Yant  (A.S.H.V.E.  TRANSACTIONS,  Vol.  35, 
1929). 

657 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

increase  can  be  compensated  for  by  raising  the  MRT.  Similarly,  with  a 
higher  air  temperature  the  same  total  heat  loss  will  be  maintained  by  a 
correspondingly  lower  MRT. 

The  loss  by  evaporation  depends  on  the  air  temperature,  air  movement, 
and  humidity;  it  is  increased  if  the  humidity  is  reduced.  For  the  usual 
conditions  of  heating  by  radiators  or  convectors,  where  the  air  tempera- 
ture ranges  from  70  F  to  73  F,  approximately  75  per  cent  of  the  total  heat 
loss  of  400  Btu  per  hour  occurs  by  radiation  and  convection,  and  the 
balance,  or  100  Btu  per  hour,  occurs  by  evaporation.  In  the  case  of 
radiant  heating,  if  the  air  temperature  is  reduced  to  60  F,  84  per  cent  of 
the  400  Btu  per  hour,  or  336  Btu  per  hour,  is  lost  by  radiation  and  con- 
vection, and  64  Btu  per  hour  are  lost  by  evaporation. 

The  mean  normal  surface  temperature  of  the  human  body,  taken  over 
the  whole  area,  including  not  only  the  exposed  skin  surface  but  also  sur- 
faces of  the  clothes  and  the  hair,  has  been  very  extensively  used  as  75  F, 
particularly  in  British  literature.  However,  results  obtained  by  Aidrich2 
in  rooms  in  which  the  air  and  wall  surface  temperatures  were  approxi- 
mately 72  F  gave  mean  values  nearer  to  83  F  than  to  75  F. 

The  mean  body  surface  temperature  which  will  maintain  the  optimum 
heat  loss  by  radiation  and  convection  in  a  uniform  environment  of  72  F 
may  be  calculated  from  fundamental  equations  for  radiation  and  natural 
convection  by  substituting  a  comparable  cylinder  for  the  body.  Heilman3 
gives  the  following  equations: 

-(ro)']  « 

('-.-r.r> 

where 

HT  =*  heat  loss  by  radiation,  Btu  per  square  foot  per  hour. 
He  »  heat  loss  by  convection,  Btu  per  square  foot  per  hour. 

Zs  =  absolute  temperature  of  the  body  surface,  degrees  Fahrenheit. 
Tw  -  absolute  temperature  of  the  walls,  degrees  Fahrenheit. 

TV  =  absolute  temperature  of  the  air,  degrees  Fahrenheit. 


m  _  „    __ 

D  «  diameter  of  cylinder,  inches. 
e  «  the  ratio  of  actual  emission  to  black  body  emission, 

If  it  be  assumed  that  a  normal  adult  has  an  average  height  of  5  ft  8  in. 
and  an  average  body  surface  area  of  19.5  sq  ft,  the  surface  of  his  body  will 
have  the  same  area  as  that  of  a  cylinder  5  ft  8  in.  long  with  a  diameter  of 
13.15  in*  The  value  of  e  for  skin  and  clothing  is  practically  0.95.  Ta  and 
Tw  are  each  taken  as  72  F,  or  532  Absolute.  The  surn^  of  HT  and  Hc  is 
taken  to  be  15.4  Btu  per  square  foot  per  hour,  which  is  derived  as  the 
normal  rate  of  heat  loss  due  to  convection  and  radiation  from  a  sedentary 
individual  by  dividing  his  total  sensible  heat  loss  by  his  area.  Solution  of 


*A  study  of  Body  Radiation,  by  L.  B.  Aidrich  (Smithsonian  Miscellaneous  Collections,  Vol,  81,  No.  6, 
December,  1928). 

•Surface  Heat  Transmission,  by  R.  H.  HeUman  (Trans.  A*S,M.E.V  F#ds  and  Steam  Power 
Vol.  51,  No.  22,  September-December,  1929). 


CHAPTER  38 — RADIANT  HEATING 


Equations  1  and  2,  using  average  figures  as  outlined,  gives  a  value  of 
approximately  83  F  for  the  normal  temperature  of  the  body  surface. 
This  agrees  more  closely  with  the  values  obtained  by  Aldrich  than  with 
the  75  F  used  by  British  investigators. 

British  Equivalent  Temperature 

The  British  Equivalent  Temperature  BET  is  the  temperature  of  an 
environment  which  is  effective  in  controlling  the  rate  of  sensible  heat  loss 
from  a  sizable  black  body  in  still  air  when  the  body  has  a  maintained 
surface  temperature  of  83  F.  The  BET  is,  therefore,  a  function  of  both 
the  air  temperature  and  the  mean  radiant  temperature.  Its  numerical 
value  in  a  uniform  environment  (walls  and  air  at  the  same  temperature) 
is  equal  to  the  temperature  of  the  walls  and  air.  In  a  non-uniform  environ- 
ment (walls  and  air  at  different  temperatures)  the  BET  is  equivalent  to 
that  of  a  uniform  environment  in  which  an  83  F  surface  loses  sensible 
heat  at  the  same  rate  as  it  does  in  the  non-uniform  environment.  As 
originally  defined,  the  BET  was  based  on  a  body  surface  temperature  of 
75  F,  but  83  F  has  been  accepted  as  giving  results  more  nearly  conforming 
with  American  practice4.  The  higher  the  BET  the  less  the  heat  loss  from 
the  body,  the  rate  of  loss  in  still  air  being  approximately  proportional  to 
the  difference  between  the  BET  and  the  mean  body  surface  temperature. 

If  the  BET  were  83  F,  there  could  be  no  sensible  heat  loss  from  a 
surface  at  that  temperature,  so  the  temperature  of  a  normal  body  surface 
would  have  to  rise  to  a  point  where  the  heat  generated  in  the  tissues  could 
be  dissipated. 

When  convected  heat  is  used,  the  temperatures  of  the  air  and  walls  are 
nearly  the  same,  and  the  optimum  value  of  the  BET  from  the  physio- 
logical point  of  view  is  72  F.  Under  these  conditions  the  mean  surface 
temperature  of  a  normal  body  would  have  the  optimum  value  of  83  F 
because  the  rate  of  heat  loss  by  radiation  and  convection  would  be  15.4 
Btu  per  square  foot  per  hour  and  that  by  evaporation  5.1  Btu  per  square 
foot  per  hour,  which  would  just  balance  the  rate  of  heat  production  of 
20.5  Btu  per  square  foot  per  hour.  This  BET  of  72  F  in  a  uniform 
environment  is  exactly  equivalent  to  the  effective  temperature  of  66  F  as 
defined  by  the  AMERICAN  SOCIETY  OF  HEATING  AND  VENTILATING 
ENGINEERS  (see  Chapter  2),  because,  in  a  uniform  environment,  a  dry- 
bulb  temperature  of  72  F  in  still  air  with  a  relative  humidity  of  30  per  cent 
gives  an  effective  temperature  of  66  F,  which  has  been  determined  to  be 
the  optimum. 

METHODS  OF  APPLICATION 

There  are  two  general  methods  of  applying  radiant  heating,  as  follow: 

1,  By  warming  the  interior  surfaces  of  the  building.  Pipe  coils  are  embedded  in  the 
concrete  or  plaster  of  the  walls,  ceiling  or  floors,  the  heating  medium  being  hot  water  or, 
in  some  cases,  steam.  This  has  the  effect  of  warming  the  entire  concrete  or  plaster 
surface  in  which  the  pipes  are  embedded.  Since  the  temperature  of  the  heating  medium 
should  not  exceed  about  120  F  on  account  of  the  possibility  of  cracking  the  plaster,  the 


^Application  of  the  Eupatfceoscope  for  Measuring  the  Performance  of  Direct  Radiators  and  Convectors 
in  Terms  qf  Equivalent  Temperatures,  by  A.  C.  Willard,  A,  P,  tCrat?,  and  M.  K.  Fahnestock  (A.S.H.V.E, 
Journal,  Beating,  Piping  and  Air  Conditioning,  July,  1933), 

6S9 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

area  of  the  panel  must  be  sufficient  to  supply  the  requisite  quantity  of  heat  at  this  low 
temperature.  When  carefully  designed,  this  method  produces  comfortable  and  eco- 
nomical results. 

2.  By  attaching  separate  heated  plates  or  panels  to  the  interior  surfaces  of  the  structure. 
These  plates  or  panels  are  placed  either  in  an  insulated  recess  flush  with  the  surface  of 
the  walls  or  ceiling  or  bolted  on  its  face.  They  may  be  decorated  as  desired.  As  it  is 
difficult  to  make  an  invisible  joint  between  the  edge  of  such  a  plate  and  the  plaster,  it  is 
common  to  use  a  frame  of  plaster,  wood,  metal  or  composition  around  the  panel.  These 
plates  may  be  placed  either  on  the  ceiling  or  the  wall,  or  in  some  cases  as  a  margin 
around  the  edge  of  the  floor.  If  floor  heating  is  required  the  temperature  over  the  whole 
area  should  not  exceed  70  F. 

If  the  entire  warm  surface  is  installed  at  one  end  of  the  room  there  may 
be  a  marked  difference  between  the  BET  on  the  two  sides  of  a  body  in  the 
room.  It  is  usually  desirable  therefore  that  the  heat  be  distributed  at 
different  points  in  the  room  so  that  no  uncomfortable  effects  will  be  felt 
from  unequal  heating. 

PRINCIPLES  OF  CALCULATION 

The  calculations  for  radiant  heating  are  entirely  different  from  those 
for  convective  heating.  The  purpose  of  the  latter  is  to  determine  the  rate 
of  heat  loss  from  the  room  by  conduction,  convection,  and  radiation  when 
maintained  in  the  desired  condition;  radiant  heating  involves  the  regu- 
lation of  the  rate  of  heat  loss  per  square  foot  from  the  human  body. 

The  first  step  in  the  calculations  for  radiant  heating  is  to  ascertain  the 
necessary  mean  radiant  temperature  (MRT) ;  next,  the  size,  temperature, 
and  disposition  of  the  heating  surfaces  required  in  the  room  to  produce 
this  MRT  are  estimated;  and  after  this  the  determination  of  the  convec- 
tive heat  is  made. 

Mean  Radiant  Temperature 

If  the  whole  of  the  interior  surface  of  a  room  were  at  the  same  tempera- 
ture, this  *  temperature  would  represent  the  MRT.  Such  a  condition 
seldom  exists,  however,  since  the  actual  surface  temperature  in  any 
heated  space  having  surfaces  exposed  to  the  outer  air  varies  greatly  for 
different  sides  of  the  enclosure.  It  is  therefore  necessary  to  ascertain  by 
calculation  the  mean  of  these  interior  surface  temperatures. 

The  mean  temperature  in  this  sense  is  not  the  arithmetic  average  of  the 
actual  thermometric  temperatures  of  the  surfaces,  but  the  temperature 
corresponding  to  the  average  rate  of  heat  emission  per  square  foot  of 
surface.  The  temperature  corresponding  to  this  mean  emission  can  be 
taken  from  Table  L  Conversely,  the  emission  at  different  temperatures 
and  also  the  emissivity  factors  can  be  obtained  from  this  table.  For 
instance,  1  sq  ft  of  surface  at  50  F  will  emit  104.9  Btu  per  square  foot  per 
hour  to  surroundings  at  absolute  zero  if  the  emissivity  of  the  surface  is  0.9. 

If  the  area  in  square  feet  of  each  part  of  the  space  is  multiplied  by  the 
emission  value  corresponding  to  its  actual  temperature,  and  these  products 
are  added  together,  the  gross  amount  of  radiant  heat  discharged  into  the 
room  by  the  wall  surface  per  hour  is  obtained.  This  quantity,  divided  by 
the  total  interior  surface,  gives  the  average  amount  of  heat  coming  into 
the  room  from  the  surface  of  the  walls  per  square  foot  of  surface  per  hour. 

Interpolating  in  Table  1,  the  total  radiation  from  a  surface  at  83  F  for 

660 


CHAPTER  38 — RADIANT  HEATING 


TABLE  1.    TOTAL  BLACK  BODY  RADIATION  TO  SURROUNDINGS  AT  ABSOLUTE  ZEROa 


BODY 

OR 

MEAN 
RADIANT 
TEMPER- 
ATURE 
Deg 
Fahr 

Radiation  in  Btu  per  square  foot  per  hour 
emitted  to  surroundings  with  a  tempera- 
ture of  absolute  zero  by  bodies  at  various 
temperatures  and  with  emissivity  factor  e 

BODY 

OR 

MEAN 
RADIANT 
TEMPER- 
ATURE 
Deg 
Fahr 

Radiation  in  Btu  per  square  foot  per  hour  emitted 
to  surroundings  with  a  temperature  of  absolute 
zero   by   bodies   at   various   temperatures   and 
with  emiesivity  factor  e 

€ 

1.00 

0.9S 

0.90 

0.80 

e 
1.00 

0.95 

0.90 

0.80 

30 

99.3 

94.3 

89.4 

79.4 

71 

136.5 

129.6 

122.9 

109.3 

35 

103.5 

98.3 

93.2 

82.8 

72 

137.4 

130.5 

123.6 

109.9 

40 

107.6 

102.4 

96.8 

86.1 

73 

138.4 

131.5 

124.5 

110.6 

45 

112.1 

106.5 

100.9 

89.7 

74 

139.6 

132.6 

125.6 

111.7 

46 

112.9 

107.3 

101.6 

90.4 

75 

141.0 

133.9 

126.9 

112.8 

47 

113.9 

108.2 

102.5 

91.1 

80 

146.6 

139.4 

132.0 

117.4 

48 

114.8 

109.1 

103.4 

91.9 

85 

152.3 

144.6 

137.1 

121.9 

49 

115.6 

109.9 

104.1 

92.4 

90 

157.9 

149.9 

142.1 

126.4 

50 

116.5 

110.6 

104.9 

93.2 

100 

169.6 

161.1 

152.6 

135.7 

51 

117.5 

111.6 

105.8 

94.0 

110 

181.6 

172.5 

163.5 

145.4 

52 

118.4 

112.5 

106.5 

94.7 

120 

194.8 

185.0 

175.4 

155.9 

53 

119.4 

113.4 

107.4 

95.5 

130 

210.1 

199.6 

189.1 

168.1 

54 

120.2 

114.2 

108.2 

96.2 

140 

223.2 

212.1 

201.0 

178.5 

55 

121.1 

115.1 

109.0 

96.9 

150 

237.1 

225.2 

213.5 

189.7 

56 

122.1 

116.0 

109.9 

97.7 

160 

251.1 

238.8 

226.0 

201.0 

57 

123.1 

117.0 

110.9 

98.5 

170 

270.5 

257.0 

243.5 

216.4 

58 

124.0 

117.8 

111.6 

99.2 

180 

288.0 

273.8 

259.1 

230.4 

59 

124.9 

118.6 

112.4 

99.9 

190 

306.5 

291.0 

275.8 

245.1 

60 

125.8 

119.5 

113.4 

100.7 

200 

325.2 

309.0 

292.8 

260.3 

61 

126.6 

120.3 

114.0 

101.4 

210 

348.0 

330.6 

313.1 

278.4 

62 

127.7 

121.4 

114.9 

102.2 

220 

371.5 

353.0 

334.4 

297.1 

63 

128.6 

122.2 

115.8 

102.9 

250 

437.8 

415.9 

394.0 

350.2 

64 

129.6 

123'.  1 

116.7 

103.7 

300 

575.0 

546.1 

517.5 

460.0 

65 

130.5 

124.0 

117.5 

104.4 

350 

740.0 

703.0 

666.0 

592.0 

66 

131.6 

125.0 

118.4 

105.4 

400 

942.1 

895.0 

847.5 

753.5 

67 

132.5 

125.9 

119.3 

106.0 

450 

1176.0 

1117.0 

1059.0 

941,0 

68 

133.5 

126.8 

120.1 

106.8 

500 

1464.0 

1390.0 

1318.0 

1171.0 

69 

134.5 

127.8 

121.1 

107.6 

550 

1791.  D 

1701.0 

1613.0 

1434.0 

70 

135.5 

128.8 

121.9 

108.4 

600 

2405.0 

2284.0 

2165.0 

1925.0 

aThese  factors  are  calculated  from  the  formula 


0.1723  X  T*  \ 
,  100,000,000  ) 

where 

Q  ••  total  black  body  radiation,  Btu  per  square  foot  per  hour. 
e  •"  emisaivity, 
T  «*  absolute  temperature,  degrees  Fahrenheit. 

an  emissivity  of  0.95  is  142  Btu  per  square  foot  per  hour.  The  difference 
between  142  Btu  and  the  average  amount  of  heat  coming  into  the  room  is 
the  amount  which  will  be  lost  per  square  foot  per  hour  by  radiation  from 
a  body  at  83  F.  If  a  rate  at  which  it  is  desired  that  heat  be  lost  from  the 
body  by  radiation  and  convection  be  assumed,  the  mean  radiant  emission 
from  the  walls  required  to  give  the  desired  result  can  be  determined  from 
Table.  1,  as  can  also  the  required  air  temperature  for  the  corresponding 
convective  effect, 

The  determination  of  the  amount  of  radiant  heating  surface  needed  in 
a  room  requires  knowledge  of  the  climate,  the  type  of  structure,  the  type 
of  heating,  and  the  surface  temperature  of  the  walls.  This  problem  can 
be  solved  only  on  an  empirical  basis.  After  some  experience,  however, 

661 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

it  is  possible  to  estimate  these  variables  with  a  considerable  degree  of 
accuracy  for  any  climate  or  construction. 

Assume  that  a  mean  radiant  temperature  of  65  F  is  desired.  Table  1 
shows  that  with  all  the  walls  at  this  temperature,  and  with  an  emissivity 
of  0.95,  the  gross  heat  emission  is  124  Btu  per  square  foot  per  hour.  The 
total  emission  of  radiation  into  the  room  from  that  surface  would  there- 
fore be  A  X  124,  where  A  is  the  total  inside  area  of  the  room.  This  is  the 
desired  emission. 

If  the  whole  area  be  divided  into  a  number  of  different  parts  which  are 
each  at  a  uniform  temperature — ai,  as,  as, — and  each  is  multiplied  by  the 
value  of  the  heat  emission  corresponding  to  that  temperature,  and  if  all 
these  products  are  added  together,  their  sum  will  represent  the  total 
actual  emission  of  radiation  into  the  room  at  these  temperatures  without 
the  aid  of  any  hot  surface. 

The  difference  between  the  desired  emission  and  the  actual  emission 
represents  the  additional  heat  which  must  be  supplied  by  the  hot  surface. 
The  temperature  of  the  proposed  hot  surface  must  then  be  selected,  and 
its  emission  per  square  foot  at  that  temperature  determined  from  Table  1. 
This  emission  is  divided  into  the  additional  amount  of  heat  needed,  ad- 
justed for  the  fact  that  the  heating  units  will  shield  the  walls  behind 
them,  and  the  quotient  obtained  will  be  the  area  of  the  required  heating 
surface. 

It  is  evident  that  this  method  of  calculation  is  approximate,  and 
depends  for  its  accuracy  on  a  correct  estimate  of  the  ultimate  surface 
temperatures  attained  by  the  actual  wall  surfaces. 

It  is  necessary  also  to  calculate  how  much  heat  will  be  given  off  by  the 
same  surfaces  by  convection,  and  thereby  to  determine  ^ whether  this 
amount  of  convected  heat  will  warm  entering  ventilating  air  to  the  tem- 
perature maintained.  If  it  will  not,  additional  convection  surfaces  must 
be  introduced  to  make  up  the  deficiency. 

MEASUREMENT  OF  RADIANT  HEATING 

Convection  heating,  having  as  its  object  the  raising  of  the  air  tempera- 
ture to  a  specified  degree,  must  be  measured  by  thermometric  methods 
which  indicate  essentially  the  air  temperature,  and  not  the  rate  of  heat 
loss  from  the  human  body.  Radiant  heating,  having  as  its  object  the 
control  of  the  rate  of  heat  loss  from  the  human  body,  ^  can  be  measured 
only  by  methods  which  basically  are  calorirnetric,  that  is,  which  measure 
directly  the  rate  of  heat  loss  from  an  object  maintained  at  the  temperature 
of  the  body,  irrespective  of  air  temperature. 

The  apparatus  for  this  purpose  consists  essentially  of  a  hollow  sphere, 
or  cylinder,  containing  a  fluid  which  can  be  maintained  accurately  at  83  F 
(the  accepted  mean  surface  temperature  of  the  human  Jbody),  with  an 
accurate  means  of  measuring  the  rate  of  heat  supply  required  to  maintain 
the  temperature  at  that  exact  point.  The  latter  measurement  can  be 
made  with  sufficient  accuracy  by  electrical  methods,  Although  a  BET  of 
72  F  is  desirable,  the  mean  radiant  and  air  temperatures  may  both  vary, 
provided  the  heat  loss  by  radiation  and  convection  from  a  surface  at 
83  F  is  maintained  at  the  rate  of  15.4  Btu  per  square  foot  per  hour, 

662 


CHAPTER  38 — RADIANT  HEATING 


15  4 

which  corresponds  to  O^TK   =   ^.5  watts  per  square  foot  of  exposed 


surface. 

This  instrument,  the  eupatheoscope,  can  readily  be  adapted  as  a  thermo- 
stat by  electrical  control  to  shut  off  or  turn  on  heat  when  the  critical 
temperature  of  83  F  in  the  vessel  is  increased  or  decreased.  A  modifi- 
cation of  the  instrument  is  called  the  eupatheostat. 

Another  instrument  for  maintaining  comfort  conditions  is  at  present 
available  only  in  a  model  adapted  to  British  practice  as  it  is  designed  for  a 
temperature  of  75  F.  It  consists  of  a  blackened  copper  sphere  of  approxi- 
mately 6  in.  diameter  in  which  is  housed  a  cylindrical  sump  containing  a 
volatile  liquid.  In  operation,  a  small  electric  heating  coil  drawing  about 
5  watts  creates  in  the  sphere  a  vapor  pressure  which  is  constant  as  long  as 
the  heat  losses  from  the  sphere  are  standard.  If  the  temperature  of  the 
air  or  the  MRT  becomes  too  high  for  comfort,  a  greater  pressure  is 
created,  owing  to  a  smaller  loss  of  heat  from  the  sphere.  This  increase  of 
pressure  acts  on  a  diaphragm  and  shuts  off  the  supply  of  heat  to  the  room. 

For  testing  work,  the  globe  thermometer  is  a  very  useful  instrument.  It 
consists  of  an  ordinary  mercury  thermometer,  with  its  bulb  placed  in  the 
center  of  a  sphere  from  6  in.  to  9  in.  in  diameter,  usually  made  of  thin 
copper  and  painted  black.  The  temperature  thus  recorded  is  termed  the 
radiation-convection  temperature  . 

EXAMPLE 

Example  L  The  surface  areas,  temperatures,  and  emissions  for  a  room  having  a 
volume  of  5760  cu  ft  are  given  in  Table  2.  The  figures  for  temperatures  are  fairly 
representative  of  American  practice  with  well-built  walls,  and  are  based  on  an  emissivity 
of  0,95  which  approximates  that  of  most  paints  and  building  materials. 

TABLE  2.   SURFACE  AREAS,  TEMPERATURES,  AND  EMISSIONS  FOR  A  ROOM  OF  5760  Cu  FT 


ARBJA 
SQFT 

ASSUMED  SURFACE 
TBMPHRAXUKB 
(DJSG  FAHR) 

HIAT  EMISSION 
(BTtr  PER  SQ  FT 
PER  HOTO) 

TOTAL  HEAT  EMISSION 
FROM  AREA 
(B<pu  PBH  HOUR) 

External  Wall  

297 

50 

110.6 

32,850 

Glass                        

279 

45 

106.5 

29,710 

Inner  Wall  

480 

55 

115.1 

55,250 

Ceiling.....  „  

480 

55 

115.1 

55,250 

Floor  

480 

55 

115.1 

55,250 

Total,..,  

2016 

228,310 

The  mean  radiant  temperature  of  the  room  is 


228,310 
2016 


113.2  Btu  per  square  foot 


per  hour  which,  as  seen  from  Table  1,  corresponds  to  an  MRT  of  53  F  for  an  average 
emissivity  of  0.95* 

For  an  average  individual  having  a  body  surface  of  19.5  sq  ft,  under  conditions  of 
comfort  with  a  body  surface  temperature  of  83  F,  the  heat  given  off  by  radiation  may  be 
determined  by  means  of  Equation  1  as  217  Btu  per  hour,  or  11.1  Btu  per  square  foot  per 
hour.  This  corresponds  to  an  environmental  emission  of  142  —  11.1  »  130.9  Btu  per 
square  foot  per  hour,  and,  according  to  Table  1,  to  an  MRT  of  72  F. 

If  this  body  be  placed  in  the  room  described,  it  will  lose  heat  at  the  rate  of  19.5 
(142  —  113.2)  «  562  Btu  per  hour.  This  loss  is  345  Btu  per  hour,  or  17.7  Btu  per  square 
loot  pet  hour,  more  than  the  rate  of  heat  loss  for  comfort,  wfyich  is  only  19.5  (142  —  130.9) 
«  217  Btu  per  hour. 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

In  order  to  determine  the  amount  of  radiating  surface  necessary  to  maintain  the  MRT 
at  72  Fr  assume  the  surface  temperature  of  the  hot  plates  to  be  installed  to  be  200  F, 
which  is  approximately  the  temperature  they  would  have  if  heated  by  steam. 

The  2016  sq  ft  total  area  of  the  surfaces  of  the  room  multiplied  by  130.9,  which  is  the 
emission  in  Btu  per  square  foot  per  hour  necessary  to  maintain  a  body  surface  tempera- 
ture of  83  F,  gives  a  total  desired  emission  of  263,890  Btu  per  hour.  It  is  necessary  to 
supply  enough  radiant  heating  surface  to  increase  the  total  actual  mean  radiant  heat 
emission  by  the  room  from  228,310,  as  shown  in  Table  2,  to  the  263,890  Btu  desired. 
The  additional  heat  needed  is  the  difference  between  these  figures,  or  35,580  Btu.  Since, 
from  Table  1,  the  emission  per  square  foot  at  200  F  is  309  Btu,  the  required  radiant 

o  K     c  QA 

heating  surface  needed  is      'nn     =  115  square  feet.   The  effect  of  this  surface  suitably 
ouy 

placed  would  be  to  raise  immediately  the  mean  radiant  temperature  to  the  required 
degree  and  to  maintain  it  at  that  value  as  long  as  the  surfaces  remained  at  the  values 
assumed. 

In  the  solution  of  this  particular  example,  the  radiation  loss  from  the 
human  body  was  selected  as  217  Btu  per  hour,  which  is  that  taking  place 
under  optimum  comfort  conditions,  with  a  body  surface  temperature  of 
83  F  in  a  uniform  environment  at  72  F.  The  mean  radiant  temperature 
necessarily  was  72  F.  If  the  optimum  BET  of  72  deg  Fahr  is  desired, 
an  air  temperature  of  72  F  also  must  be  maintained.  If  it  is  desired  to 
maintain  a  lower  air  temperature  than  this,  a  mean  radiant  temperature 
greater  than  72  F  must  be  selected  and  the  radiation  loss  from  the  in- 
dividual must  be  recalculated  from  Equation  1. 

The  calculation  may  be  simplified  by  preparing  tables  showing,  at  the 
usual  temperatures,  the  area  of  hot  surface  required  to  bring  each  square 
foot  of  actual  wall  surface  at  various  temperatures  up  to  a  general 
standard  of  from  60  F  to  70  F.  It  would  then  be  necessary  only  to 
multiply  the  respective  areas  by  the  appropriate  factors,  and  to  add  the 
results,  to  obtain  the  required  total. 

REFERENCES 

Room  Warming  by  Radiation,  by  A.  H.  Barker  (A.S.H.V.E.  TRANSACTIONS,  Vol.  38, 
1932). 

Panel  Warming,  by  L.  J.  Fowler  (A.S.H.V.E.  TRANSACTIONS,  VoL  36,  1930). 

Calculations  for  Radiant  Heating,  by  T.  Napier  Adlam  (Heating  and  Ventilating, 
October,  1931). 

Principles  of  Calculation  of  Low  Temperature  Radiant  Heating,  by  A.  H.  Barker 
(Proceedings  of  The  Institution  of  Heating  and  Ventilating  Engineers,  London,  VoL  30, 
1931). 

Application  of  the  Eupatheoscope  for  Measuring  the  Performance  of  Direct  Radiators 
and  Convectors  in  Terms  of  Equivalent  Temperatures,  by  A.  C.  Willard,  A.  P,  Kratz 
and  M.  K-  Fahnestock  (A.S.H.V.E.  Journal  Section,  Heating,  Piping  and  Air  Con~ 
ditioning,  July,  1933). 

What  will  be  the  Future  Development  of  Heating  and  Air  Conditioning,  by  W.  H. 
Carrier  (Heating,  Piping  and  Air  Conditioning,  January,  1933). 

Method  of  Installing  the  Panel  Heating  System  in  the  British  Embassy  Building 
(Heating,  Piping  and  Air  Conditioning,  July,  1934). 

Panel  Heating,  by  C.  M.  Gates  (Proceedings  of  Institution  of  Heating  and  Ventilating 
Engineers,  London,  Vol.  30,  1931). 

Notes  on  Electric  Warming  with  Special  Reference  to  Low  Temperature  Panel 
Systems,  by  R.  Grierson  (Proceedings  of  Institution  of  Heating  and  Ventilating  Engineers, 
London,  Vol.  28,  1929). 

Radiant  Heat,*by  A.  F.  Dufton  (Proceedings  of  Institution  of  Heating  and  Ventilating 
Engineers,  London,  VoL  30,  1931). 

Radiant  Heat,  by  A.  F.  Dufton  (Proceedings  of  Institution  of  Heating  and  Ventilating 
Engineers,  London,  VoL  31,  1932). 

Notes  on  the  Theory  of  Radiant  Heating,  by  C,  G.  Heys  Hallett  (Proceedings  of 
Institution  of  Heating  and  Ventilating  Engineers,  London,  VoL  29,  1930). 

664 


CHAPTER  38 — RADIANT  HEATING 


PROBLEMS  IN  PRACTICE 

1  •  Name  three  ways  that  heat  is  lost  from  the  human  body. 

By  radiation,  convection,  and  evaporation. 

2  •  What  is  the  mean  normal  surface  temperature  of  the  human  body  as 
determined  for  the  United  States? 

83  F. 

3  •  What  is  the  exact  purpose  of  radiant  heating? 

Radiant  heating  regulates  the  heat  loss  from  the  human  body. 

4  •  How  is  the  required  amount  of  radiant  heating  surface  found? 

By  calculating  the  desired  emission  and  the  actual  emission,  and  finding  their  difference. 
This  is  additional  heat  which  must  be  supplied  by  the  hot  surface. 

5  •  After  finding  the  required  heat,  how  is  the  necessary  hot  surface  area 
calculated? 

Having  selected  the  hot  surface  temperature,  find  the  emission  per  square  foot  from 
Table  1.  This  rate  divided  into  the  heat  required  gives  the  area  of  the  necessary  heating 
surface. 

6  •  Is  the  heat  generated  in  the  body  affected  by  action?     If  so,  does  it  vary 
greatly? 

Yes.  With  hard  work  or  energetic  exercise,  the  total  heat  generated  in  the  body  may  be 
five  to  six  times  that  generated  when  it  is  at  rest. 

7  •  When  and  why  does  the  human  body  feel  cold? 

The  body  feels  cold  not  only  when  it  loses  heat  at  a  greater  rate  than  it  can  generate  it, 
but  also  when  heat  is  abstracted  from  the  body  disproportionately.  The  human  body 
does  not  require  any  heat  from  without  because  it  generates  more  heat  than  is  sufficient 
to  maintain  the  correct  temperature;  therefore,  it  is  only  necessary  to  provide  conditions 
that  will  maintain  the  correct  ratio  of  losses. 

8  •  a.  Where  did  radiant  heating  derive  its  name? 

b.  What  is  actually  meant  by  radiant  heating? 

a.  The  term  radiant  heaters  was  introduced  about  25  years  ago  to  designate  flat  heating 
surfaces  made  to  give  off  practically  all  their  heat  by  radiant  ether  waves  instead  of 
relying  on  convected  warm  air. 

b.  The  term  radiant  heating  now  applies  to  methods  of  heating  where,  instead  of  heating 
the  air  to  a  predetermined  temperature,  flat  heating  surfaces  are  so  placed  in  a  room 
that  the  average  virtual  temperature  of  all  wall,  ceiling,  floor,  and  glass  surfaces 
exposed  to  the  body  is  just  sufficient  to  prevent  the  body's  losing  too  much  heat  by 
radiation.    The  air  temperature  can  be  much  cooler  with  radiant  heating  because 
radiation  losses  from  the  body  are  compensated. 

9  •  What  kind  of  heating  surfaces  are  in  general  use? 

The  heating  units  may  have  flat  iron  surfaces  heated  with  steam  and  placed  under 
windows,  or  hot  water  pipes  may  be  embedded  in  the  floor,  walls,  or  ceilings.  Electrical 
radiant  heaters  are  made  by  embedding  resistance  elements  in  porcelain  or  electric 
conductors  woven  into  a  thick  paper  which  can  be  fastened  to  the  walls  or  ceilings. 

10  •  What  kind  of  heat  rays  are  commonly  generated  in  radiant  heating? 
Give  examples. 

All  heat  rays  are  generally  assumed  to  be  the  same  as  light  rays;  they  travel  at  the  speed 
of  light,  but  they  are  invisible  and  loneer.  The  rays  used  in  heating  are  0.00005  to 
0.0001  in.  long  compared  with  visible  red  rays  of  about  0,000027  in. 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

11  •  What  natural  evidence  have  we  that  air  temperature  alone  is  no  criterion 
of  comfort  and  that  radiant  heat  affects  the  body  more  quickly? 

When  standing  in  the  sunshine  on  a  cool  spring  day,  a  person  feels  perfectly  comfortable, 
but  when  a  cloud  passes  over  the  sun,  he  instantly  feels  much  cooler  as  the  shadow 
reaches  him.  A  shielded  thermometer  recording  the  temperature  of  the  air  shows  no 
reduction  in  air  temperature  in  so  short  a  period,  so  that  the  person  actually  feels  a 
sensation  of  cold  which  an  ordinary  thermometer  cannot  register.  This  shows  that 
light  and  heat  rays  are  shut  off  simultaneously  and  travel  at  the  same  speed;  it  also 
proves  that  radiant  rays  affect  the  comfort  of  the  body  quicker  than  air  temperature 
does. 


666 


Chapter  39 

ELECTRICAL  HEATING 

Resistors,    Heating    Elements,    Electric    Heaters,    Unit    Heaters, 

Central  Fan  Heating,  Electric  Steam  Heating,  Electric  Hot  Water 

Heating,    Heat    Pump,    Control,    Calculating    Capacities,    Power 

Problems,  Electric  Heating  Data 

TTLECTRIC  heating  has  a  logical  and  a  rapidly  growing  place  in  the 
XL/  heating  industry  because  of  its  advantages  of  flexibility,  cleanliness, 
safety,  convenience,  and  ease  of  control.  Electric  heating  practice  has 
many  basic  principles  in  common  with  fuel  heating,  but  there  are  also 
important  differences.  The  advantages  of  good  building  insulation  are 
even  more  important  in  electric  heating  than  for  fuel  heating,  because  the 
initial  cost  per  Btu  is  usually  higher. 

All  heat  is  a  form  of  energy.  Fuels  hold  stored  chemical  energy  which 
is  released  into  heat  by  combustion.  Electrical  power  is  a  form  of  energy 
which  can  be  released  into  heat  by  passing  it  through  a  resisting  material. 
Both  fuel  and  electric  heating  have  two  divisions:  first,  the  conversion  of 
energy  into  heat;  second,  the  distribution  and  practical  use  of  the  heat 
after  it  is  produced. 

In  converting  the  chemical  energy  of  fuels  into  heat  by  combustion, 
there  is  necessarily  a  considerable  variation  in  thermal  efficiency.  This 
is  not  true,  however,  when  converting  electric  power  into  heat,  because 
100  per  cent  of  the  energy  applied  in  the  resistor  is  always  transformed 
into  heat.  In  electric  heating  practice  the  engineer  need  not  be  concerned 
about  efficiencies  of  heat  production,  but  rather  about  efficiencies  of  heat 
utilization. 

DEFINITIONS 

Definitions  of  terms  used  in  fuel  heating  are  given  in  Chapter  41.  The 
following  terms  apply  particularly  to  electric  heating: 

Electric  Resistor ;  A  material  used  to  produce  heat  by  passing  an  electric  current 
through  it.  * 

Electric  Heating  Elements  A  unit  assembly  consisting  of  a  resistor,  insulated 
supports,  and  terminals  for  connecting  the  resistor  to  electric  power. 

Electric  Heater:  A  complete  assembly  of  heating  elements  with  their  enclosure, 
ready  for  installation  in  service. 

RESISTORS 

Solids,  liquids^  and  gases  may  be  used  as  resistors,  but  most  com- 
mercial electric  heating  elements  have  solid  resistors,  such  as  metal 
alloys,  and  non-metallic  compounds  containing  carbon.  In  some  types  of 
electric  boilers,  water  forms  the  resistor  which  is  heated  by  an  alternating 
currerit  of  electricity  pa&stog  through  it.  One  of  the  more  common 

eer 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

resistors  is  nickel-chromium  wire  or  ribbon  which,  in  order  to  avoid 
oxidation,  contains  practically  no  iron. 

HEATING  ELEMENTS 

Commercial  electric  heating  elements  are  divided  into  open  type 
elements,  enclosed  type  elements,  and  cloth  fabrics.  Open  type  elements 
have  resistors  exposed  to  view.  The  resistors  may  be  coils  of  wire  or 
metal  ribbon,  supported  by  refractory  insulation,  or  they  may  be  non- 
metallic  rods,  mounted  on  insulators.  Open  type  elements  are  used 
extensively  for  operation  at  high  temperatures  when  radiant  heat  is 
desired.  They  are  also  frequently  used  at  low  temperatures  for  convec- 
tion and  fan  circulation  heating,  especially  in  large  installations. 

Enclosed  type  elements  have  metallic  resistors  embedded  in  a  refractory 
insulating  material,  and  encased  in  a  protective  sheath  of  metal.  Fins  or 
extended  surfaces  may  be  used  to  add  heat-dissipating  area.  Enclosed 
elements  are  made  in  many  forms,  such  as  strips,  rings,  plates,  and  tubes. 
Strip  elements  are  used  for  clamping  to  surfaces  requiring  heat  by  con- 
duction, and  in  convection  and  fan  circulation  air  heaters.  Ring  and 
plate  elements  are  used  in  electric  ranges,  waffle  irons,  and  in  many  small 
air  heaters.  Tubular  elements  may  be  immersed  in  liquids,  cast  into 
metal,  and,  when  formed  into  coils,  used  in  electric  ranges  and  air  heaters. 
Cloth  fabrics  woven  from  flexible  resistor  wires  and  asbestos  thread,  are 
used  for  many  low  temperature  purposes. 

ELECTRIC  HEATERS 

Electric  heaters  are  classified  according  to  the  manner  in  which  they 
deliver  heat  in  practical  use,  that  is,  by  conduction,  by  radiation,  or  by 
convection.  The  term  radiator  should  not  be  used  in  electric  heating, 
because  of  confusion  between  its  established  usage  in  fuel  heating  and  the 
radiant  principle  of  many  electric  heaters. 

Among  the  uses  of  conduction  electric  heaters,  which  deliver  most  of  their 
heat  by  actual  contact  with  the  object  to  be  heated,  are  aviators'  cloth- 
ing, hot  pads,  foot  warmers,  soil  heaters,  ice  melters,  and  pipe  heaters. 
Conduction  heaters  are  useful  in  conserving  and  localizing  heat  delivery 
at  definite  points.  They  are  not  suitable  for  general  air  heating. 

Radiant  electric  heaters,  which  deliver  most  of  their  heat  by  radiation, 
have  high  temperature  incandescent  heating  elements  and  reflectors  to 
concentrate  the  heat  rays  in  the  desired  directions.  The  immediate  and 
pleasant  sensation  of  warmth  which  is  caused  by  radiant  heat  makes  this 
type  desirable  for  temporary  use  where  the  heat  rays  can  fall  directly 
upon  the  body.  They  are  not  satisfactory  for  general  heating,  as  radiant 
heat  rays  do  not  warm  the  air  through  which  they  pass.  They  must 
first  be  absorbed  by  walls,  furniture,  or  other  solid  objects  which  then 
give  up  the  heat  to  the  air.  The  location  of  radiant  heaters  is  important. 
They  should  never  face  a  window  because  some  rays  would  pass  through 
the  glass  and  be  lost.  Figs.  1  and  2  show  common  types  of  portable  and 
wall-mounted  radiant  heaters. 

Convection  electric  heaters,  designed  to  induce  thermal  air  circulation, 
deliver  heat  largely  by  convection,  and  should  be  located  and  used  in 

668 


CHAPTER  39 — ELECTRICAL  HEATING 


much  the  same  manner  as  steam  and  hot  water  radiators  or  convectors. 
They  should  have  heating  elements  of  large  area,  with  moderate  surface 
temperature,  enclosed  to  give  proper  stack  effect  to  draw  cold  air  from 
the  floor  line  (Figs.  3  and  4).  The  flexibility  possible  with  electric  heating 
elements  should  discourage  the  use  of  secondary  mediums  for  heat 
transfer.  Water  and  steam  add  nothing  to  the  efficiency  of  an  electric 
heater  and  entail  expensive  construction. 

UNIT  HEATERS 

Fan  unit  electric  heaters,  having  electric  heating  elements  combined  in 
the  same  enclosure  with  a  fan  or  blower,  are  made  in  many  styles  and  are 
excellent  for  general  air  heating.  They  should  be  located  and  used  much 


FIG.  1. 


PORTABLE  RADIANT  ELECTRIC 
HEATER 


FIG.  2.    RADIANT  ELECTRIC  HEATER 
RECESSED  IN  WALL 


t 


I 


FIG.  3. 


CONVECTION  ELECTRIC  HEATER 
ON  WALL  SURFACE 


FIG.  4. 


CONVECTION  ELECTRIC  HEATER 
RECESSED  IN  WALL 


as  steam  unit  heaters.  The  warm  air  can  be  directed  toward  the  floor,  if 
desired,  to  give  a  positive  circulation  which  will  reduce  stratification  of 
air.  Small  units  which  are  free  from  radio  interference  are  used  for 
homes;  there  are  large  units  for  industrial  plants,  substations,  power 
houses,  and  pumping  stations;  portable  units  are  useful  for  temporary 
work,  such  as  drying  out  damp  rooms,  or  for  warming  rooms  during 
construction  (Figs.  5,  6,  7  and  8). 

CENTRAL  FAN  HEATING 

Central  fan  electric  heating  systems  have  electric  heating  elements  and 
fans  or  blowers  to  circulate  the  air  through  ducts,  and  in  addition  to  the 

669 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

main  heaters  at  the  fan  location,  booster  heaters  may  be  located  in  branch 
ducts.  Humidification  or  complete  air  conditioning  can  readily  be  in- 
cluded in  the  system,  in  much  the  same  manner  as  with  steam. 

In  coordinating  the  input  of  heat  energy  and  the  volume  of  air  circu- 
lation, a  basic  difference  between  electric  heating  and  steam  heating 
enters  into  the  problem.  Steam  is  approximately  a  constant-temperature 
source  of  heat  for  any  given  pressure  as  a  change  in  air  volume  flowing 
over  steam  coils  does  not  greatly  affect  the  temperature  of  the  delivered 
air.  The  amount  of  steam  condensed  (heat  input)  varies  in  proportion  to 
the  air  volume,  but  the  surface  temperature  of  the  steam  coils  remains 


FIG.  5.   SMALL  PORTABLE  FAN  UNIT 
ELECTRIC  HEATER 


FIG.  6.    LARGE  INDUSTRIAL  TYPE  PORT- 
ABLE FAN  UNIT  ELECTRIC  HEATER 


r    t     t 


FIG.  7. 


SMALL  FAN  UNIT  ELECTRIC 
HEATER 


FIG.  8,    LARGE  INDUSTRIAL  TYPE  FAN 
UNIT  ELECTRIC  HEATER 


about  the  same.  Electric  heat  is  quite  different,  being  a  constant  source 
of  energy.  If  the  volume  of  air  flow  over  electric  heating  elements  is 
changed,  and  no  change  is  made  in  the  electrical  power  connections,  there 
will  be  a  corresponding  change  in  the  temperature  of  the  air  delivered 
because  the  electrical  energy  input  remains  constant  and  the  surface 
temperature  of  the  heating  elements  will  vary  as  is  necessary  to  force  the 
air  to  accept  all  the  heat.  With  electric  heat  the  total  heat  is  constant 
unless  some  compensating  action  is  performed  by  control.  Automatic 
modulation  to  vary  the  electrical  heat  input  and  synchronize  it  properly 
with  the  air  flow  has  been  successfully  applied  to  central  fan  systems* 

ELECTRIC  STEAM  HEATING 

Electric  steam  heating  differs  from  fuel  heating  only  in  the  use  of  electric 
boilers  to  generate  steam.  Small  boilers  usually  have  heating  elements 
of  the  enclosed  metal  resistor  type  immersed  in  the  water.  Boilers  of  this 
construction  may  be  used  on  either  direct  or  alternating  current  since  the 

670 


CHAPTER  39 — ELECTRICAL  HEATING 


heat  is  delivered  to  the  water  by  contact  with  the  hot  surfaces.  To  lessen 
the  likelihood  that  the  heating  elements  will  burn  out,  they  are  made 
removable  for  cleaning  off  deposits  of  scale  which  will  restrict  the  heat 
flow.  Boilers  of  this  type  are  useful  in  industrial  plants  which  require 
limited  amounts  of  steam  for  local  processes,  and  for  sterilizers,  jacketed 
vessels,  and  pressing  machines  which  need  a  ready  supply  of  steam. 

Electric  boilers  are  entirely  automatic  and  are  well  adapted  to  inter- 
mittent operation.  It  frequently  is  economical  to  shut  down  the  main 
plant  boilers  when  the  heating  season  ends,  and  to  supply  steam  for 
summer  needs  with  small  electric  boilers  located  close  to  the  operation. 
Large  electric  boilers  are  usually  of  the  type  employing  water  as  the 
resistor.  Only  alternating  current  can  be  used,  as  direct  current  would 
cause  electrolytic  deterioration.  Large  boilers  of  this  kind  have  electrodes 
immersed  in  the  water  where  heat  is  generated  directly.  In  Canada  and 
Europe  many  successful  installations  have  been  made,  but  in  the  United 
States  the  cost  of  electric  power,  in  comparison  with  fuels,  does  not  favor 
its  general  use. 

ELECTRIC  HOT  WATER  HEATING 

Electric  hot  water  heating  offers  an  extremely  convenient  and  reliable 
means  of  supplying  all  needs  for  hot  water,  and  in  sections  of  the  country 
where  low  current  rates  have  made  it  economically  feasible,  it  enjoys 
popularity.  Electric  boilers  for  hot  water  heating  are  inexpensive, 
entirely  automatic,  and  are  insulated  to  prevent  excessive  heat  losses. 
When  lower  power  costs  can  be  secured,  by  confining  the  heating  to 
certain  fixed  hours,  water  may  be  heated  and  stored^  in  well-insulated 
tanks  for  use  when  needed.  In  large  industrial  plants  it  is  often  possible 
to  balance  power  loads  by  this  means  and  to  avoid  running  the  fuel-fired 
steam  boilers  at  night  or  over  week  ends.  In  Europe  use  has  been  made  of 
this  hot  water  storage  principle  for  heating.  Experiments  have  been 
made  in  this  country  for  heating  houses,  but  the  cost  of  serving  individual 
homes  with  the  necessary  heavy  electric  power  loads  has  proved  un- 
profitable at  rates  comparable  to  other  forms  of  heating.  The  problems 
incident  to  installing  large  storage  tanks  in  home  basements,  and  the  lack 
of  flexibility  under  variable  weather  conditions,  are  also  unfavorable 
factors. 

OIL  HEATING 

Electric  hot  oil  heating  is  useful  in  some  industrial  work  as  a  substitute 
for  superheated  steam.  Special  oil  can  be  electrically  heated  as  high  as 
600  F  and  pumped  at  a  pressure  just  sufficient  to  cause  flow.  When  used 
in  heating  coils  or  jacketed  vessels,  this  gives  a  safe,  and  convenient, 
automatic  system  for  moderate-sized  installations. 

HEAT  PUMP 

The  electric  heat  pump  is  not  strictly  an  electric  heater,  as  it  does  not 
directly  convert  electrical  power  into  heat.  It  operates  a  compressor 
electrically  which  acts  as  a  reversible  refrigerating  unit  to  ^extract  heat 
from  the  outdoor  air  in  winter  and  deliver  it  indoors  for^heating  purposes, 
and,  by  a  reversal,  to  extract  heat  from  the  indoor  air  in  summer  and 

671 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

discharge  it  outdoors.  This  system  has  been  used  in  evenly-balanced 
climates  where  the  heating  requirements  in  winter  are  about  the  same  as 
the  cooling  requirements  in  summer. 

AUXILIARY  ELECTRIC  HEATING 

In  conjunction  with  heating  systems  of  other  types,  an  auxiliary  elec- 
trical heating  arrangement  is  a  convenient  means  of  caring  for  mild  days 
in  the  spring  and  fall  which  require  little  heat  to  make  a  house  or  building- 
comfortable.  Likewise,  such  electrical  heating  might  be  used  on  ab- 
normally cold  days  to  help  out  the  main  heating  system  and  by  this  means 
reduce  the  necessary  size  of  the  system. 

Because  of  the  feeling  of  comfort  that  a  radiant  type  heater  gives, 
bathrooms  may  be  heated  electrically  with  this  type  of  heater  while  the 
rest  of  the  house  is  cared  for  by  some  other  system.  Offices  and  rooms 
which  require  heat  at  periods  when  the  main  heating  plant  is  shut  down 
are  conveniently  cared  for  electrically. 

CONTROL 

Because  the  efficiency  of  electric  heat  production  is  the  same  for  large 
or  small  units,  it  is  possible  to  reduce  heat  waste  to  a  minimum  by  apply- 
ing local  heating,  locally  controlled.  Wherever  radiant  heaters  are  used, 
thermostats  are  not  an  effective  means  of  control  and  manual  operation 
or  control  by  eupatheoscope  is  necessary.  For  all  convection  and  fan 
circulation  heaters  thermostatic  control  is  useful.  For  small  heaters 
having  ratings  up  to  about  1500  watts,  there  are  direct-acting  thermostats 
which  are  satisfactory,  but  for  larger  heaters  it  is  advisable  to  use  relays 
or  contactors,  which  should  break  all  of  the  power  lines.  All  heaters 
having  fan  circulation  should  have  the  heat  circuit  interlocked  with  the 
motor  circuit  so  that  the  fan  will  be  running  when  the  heat  is  on.  A 
thermal  fuse  or  trip  should  be  located  in  the  heat  chamber  to  throw  off 
the  heat  in  case  any  interruption  of  air  flow  should  occur;  otherwise 
undue  temperature  rise  would  result.  In  all  large  heaters  the  heating 
elements  should  be  arranged  in  groups  and  control  provided  to  vary  the 
heat  input  to  correspond  approximately  to  the  heat  demand.  If  this  is 
not  done,  and  all  the  heat  is  kept  available,  the  thermostat  will  continue 
throwing  it  on  and  off  at  short  intervals.  Except  for  central  fan  systems, 
the  heat  stages  can  be  operated  by  manual  switches,  but  automatic 
modulation  of  the  heat  load  is  usually  preferred. 

CALCULATING  CAPACITIES 

The  methods  of  calculating  heat  losses  outlined  in  Chapters  6,  7,  and  8 
may  be  used  for  electric  heating  exactly  as  for  fuel  heating.  The  total 
heat  requirements  in  Btu  per  hour  may  then  be  converted  into  the 
electrical  rating  of  an  equivalent  heating  system  by  using  the  equation  : 

Total  Btu  per  hour       ,  ,         .     ,    f       .    -  ,„. 

„..„„._;._.     __  _  kw  ratmg  of  required  electric  heating  (1) 


The  following  empirical  rules  for  estimating  electric  heater  require- 

672 


CHAPTER  39 — ELECTRICAL  HEATING 


ments  may  be  used  in  territories  where  the  heating  load  is  never  greater 
than  1500  degree  days: 

Capacity  of  heater  required  for  average  room  in  home 2  watts  per  cubic  foot. 

Capacity  of  heater  required  for  average  office  occupied 

in  the  daytime  only 1.2  watts  per  cubic  foot. 

POWER  PROBLEMS 

The  first  point  to  determine  is  the  cost  of  the  power  which  is  available 
for  electric  heating.  Unlike  fuels,  there  is  no  uniform  cost  for  electric 
power  because  of  the  unequal  cost  of  distribution  to  large  and  small 
users.  The  fact  that  electricity  cannot  be  economically  stored,  but  must 
be  used  as  fast  as  it  is  generated,  makes  it  impossible  to  operate  power 
plants  at  uniform  loads ;  hence,  even  the  time  of  use  may  affect  the  cost  of 
power. 

Homes  are  almost  universally  supplied  with  lighting  current  of  115 
volts,  which  cannot  be  used  economically  for  any  but  the  smallest  heaters. 
Usually  the  service  lines  will  not  permit  more  than  plug-in  devices.  The 
underwriters  permit  heaters  of  1250  watts  to  be  used  from  approved 
baseboard  receptacles.  Where  homes  have  230  volt  service  for  cooking 
and  water  heating,  and  rates  are  favorable,  larger  heaters  can  be  installed. 
For  industrial  purposes,  heaters  should  be  designed  to  use  polyphase 
power,  which  is  usually  supplied  at  230,  460  or  575  volts.  All  polyphase 
heaters  should  be  balanced  between  phases. 

ELECTRIC  HEATING  DATA 

Electric  heater  capacity  is  rated  in  kilowatts  (kw).  Electric  energy  is 
measured  in  kilowatt-hours  (kwhr).  Cost  of  operation  =  kw  rating 
X  hours  used  X  cost  per  kwhr. 

One  boiler  horsepower  (bhp)  —  33,471.9  Btu  per  hour, 
One  kilowatt-hour  (kwhr)       =    3,415     Btu. 

33  471  9 
One  boiler  horsepower  »     0>/I1  '     =  9.80  kwhr. 

0,4:10 

One  boiler  horsepower  will  evaporate  34.5  Ib  water  per  hour /row  and  at  212  F. 

34  5 
One  kilowatt-hour  =  Q~^~O   =  3.52  Ib  of  water  per  hour  at  212  F. 

Additional  conversion  factors  are  given  in  Chapter  41. 

PROBLEMS  IN  PRACTICE 

1  •  Why  is  electrical  energy  economically  feasible  to  use  for  certain  heating 
applications? 

a.  Because  heat  from  the  radiant  type  of  electrical  heater  is  effective  in  producing  com- 
fort for  the  occupant  almost  as  soon  as  it  is  turned  on,  the  heater  may  be  turned  off 
when  the  room  in  unoccupied. 

6.  There  is  nothing  to  freeze  in  an  electrical  heater. 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

c.  Electrical  energy  may  be  purchased  at  lower  rates  during  off-peak  periods  and  stored 
as  heat  in  water  kept  in  insulated  tanks  until  needed. 

d.  There  is  no  wasted  energy  up  the  flue  or  in  the  ash  as  with  fuel  heating. 

2  •  To  what  localities  is  electrical  heating  adaptable? 

To  those  localities  where  the  heating  season  is  relatively  mild  and  where  electrical 
energy  is  available  at  low  cost,  as  in  communities  served  by  large  hydro-electric  plants. 

3  •  Approximately  how  low  must  the  rates  be  to  permit  the  use  of  electricity 
for  heating  purposes? 

Probably  the  energy  must  sell  for  2  cents  or  less  per  kwhr.  At  2  cents  the  cost  would  be 
$5.86  per  1000  Mbh.  (See  Chapter  29  for  comparison  with  other  fuels.)  ^  This  looks 
high,  but  the  seasonal  energy  consumption  would  not  be  as  large  with  electricity  as  with 
other  fuels,  for  reasons  stated  in  Question  1. 

4  •  Why  is  automatic  control  important  in  connection  with  electrical  heating? 

The  higher  cost  of  the  energy  makes  it  essential  that  none  be  wasted. 

5  •  In  fan  heating  systems,  what  is  an  important  difference  between  a  steam 
heated  coil  and  an  electrically  heated  coil? 

A  coil  supplied  with  steam  at  constant  pressure  will  remain  at  constant  temperature 
regardless  of  •  the  amount  of  air  passing  over  it.  The  temperature  of  the  electric  coil 
supplied  with  a  constant  amount  of  energy  will  rise  if  the  air  quantity  is  decreased  and 
fall  if  the  air  quantity  is  increased. 


674 


Chapter  40 

TEST  METHODS  AND  INSTRUMENTS 

Pressure  Measurement,    Temperature  Measurement,    Air  Move- 
ment,  Humidity  Measurement,   Carbon  Dioxide  Determination, 
Dust  Determination,  Flue  Gas  Analysis,  Measurement  of  Smoke 
Density,  Heat   Transmission,  Eupatheoscope 

A  TMOSPHERIC  pressure  is  usually  measured  by  a  mercurial  barom- 
-/TL  eter  which,  in  its  simplest  form,  consists  of  a  glass  tube  about  3  ft 
long,  closed  at  the  upper  end,  filled  with  mercury  and  inverted  in  a 
shallow  bath  of  mercury.  The  pressure  of  the  atmosphere  on  the  exposed 
top  of  the  mercury  in  the  cistern  supports  a  column  of  mercury  in  the 
tube  to  a  height  of  about  30  in.  Readings  are  taken  of  the  height  of  the 
column  between  the  levels  of  mercury  in  the  tube  and  in  the  cistern. 
Atmospheric  pressure  is  the  same  as  the  pressure  exerted  by  this  supported 
column  of  mercury,  and,  in  pounds  per  square  inch,  is  equal  to  its  height 
in  inches  times  0.491,  which  is  the  weight  in  pounds  of  1  cu  in.  of  mercury. 
At  latitude  45  deg  and  sea  level,  and  at  a  temperature  of  32  F,  the  atmos- 
phere will  support  a  column  of  mercury  29.921  in.  in  height.  The  pressure 
of  14.7  Ib  per  square  inch,  derived  by  multiplying  29.921  by  0.491,  is 
called  standard  or  normal  barometric  pressure.  Since  the  height  of  the 
barometer  depends  on  the  density  of  the  mercury  as  well  as  on  the  pres- 
sure of  the  atmosphere,  and  since  the  density  is  dependent  on  the  tem- 
perature, mercurial  barometer  readings  should  always  be  corrected  for 
temperature.  An  aneroid  barometer  contains  no  liquid;  it  is  portable  but 
less  accurate  than  the  mercurial  barometer.  Atmospheric  pressure  in 
bending  the  thin  corrugated  top  of  a  partially  exhausted  metallic  box,  or 
in  distorting  a  thin-walled  bent  tube  of  metal,  is  made  to  move  a  pointer. 

Pressures  above  or  below  atmospheric  are  usually  measured  by  means 
of  gages  which  indicate  the  difference  between  the  pressure  being  measured 
and  atmospheric  pressure  at  the  same  time  and  place.  A, gage  which 
indicates  pressures  higher  than  atmospheric  is  known  as  a  pressure  gage, 
and  a  gage  which  indicates  pressures  lower  than  atmospheric  is  known  as  a 
vacuum  gage.  The  most  common  type  of  these  gages  contains  a  flexible 
hollow  brass  tube  of  oval  cross  section,  known  as  a  3  our  don  tube.  When 
subjected  to  unequal  inside  and  outside  pressures,  this  tube  tends  to 
straighten  out,  and  a  pointer  motivated  by  this  straightening  indicates 
the  pressure  difference  on  a  suitably  graduated  scale. 

High  vacuum  readings  such  as  are  encountered  in  condenser  and  steam 
jet  refrigeration  practice  are  commonly  obtained  by  the  use  of  mercury 
column  vacuum  gages.  When  the  readings  obtained  with  the  mercurial 
barometer  and  those  with  the  mercury  vacuum  gage  have  both  been 
corrected  to  32  F,  the  difference  in  the  two  readings  will  give  the  absolute 

675 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

vacuum  in  inches  of  mercury.    The  following  equation  may  be  used  to 
make  corrections  for  temperature: 

h  =  hi  [1  -  0.000101  (ti  -  t)\ 
where 

h  =  height  of  mercury  column  corrected  to  temperature  t. 
Jh  =  actual  height  of  mercury  column. 
ti  =  actual  temperature  of  mercury  column. 
t  =  temperature  to  which  column  is  to  be  corrected. 

A  gage  which  indicates  pressures  slightly  above  or  below  atmospheric  is 
known  as  a  draft  gage.  It  is  essentially  a  U  tube  containing  either  water, 
kerosene,  alcohol,  or  mercury,  with  one  leg  exposed  to  the  air  and  the 
other  connected  to  a  point  where  the  pressure  is  to  be  determined.  When 
the  pressure  being  read  is  equal  to  atmospheric,  the  level  of  the  liquid  in 
the  legs  will  be  the  same,  indicating  a  zero  gage  pressure.  When  a  pres- 
sure is  applied  to  one  leg,  one  side  will  fall  and  the  other  will  rise  an  equal 
amount.  The  difference  in  height  between  the  two  liquid  levels  indicates 
the  pressure  expressed  in  inches  of  liquid  used  in  the  gage. 

TEMPERATURE  MEASUREMENT 

In  engineering  work,  mercurial  thermometers  are  largely  employed  to 
measure  the  intensity  of  heat.  These  depend  on  the  uniform  expansion  of 
mercury  to  indicate  changes  in  temperature.  An  amount  of  mercury  held 
in  a  sealed  tube  with  a  bulb  at  one  end  will  rise  to  one  definite  level  when 
immersed  in  melting  ice,  and  to  another  definite  level  when  immersed  in 
boiling  water.  These  two  points  are  marked,  and  the  space  between  them 
is  divided  into  a  number  of  equal  portions,  each  of  which  is  called  a 
degree.  In  the  Fahrenheit  scale,  there  are  180  degrees  thus  obtained, 
while  the  centigrade  scale  has  100  and  the  Reaumur  has  80.  Like  divisions 
are  marked  off  on  the  column  above  and  below  these  two  determined 
points  in  order  that  a  greater  range  of  temperature  may  be  read. 

Thermocouples1  may  be  used  to  measure  any  range  of  temperatures  up 
to  2,900  F,  When  two  dissimilar  metals  are  joined  at  two  points  and  a 
temperature  difference  exists  between  these  junctions,  an  electromotive 
force  will  be  developed.  Its  magnitude  depends  on  the  composition  of  the 
wires  and  the  difference  in  temperature  between  the  junctions.  A  poten- 
tiometer or  sensitive  galvanometer  of  high  resistance  connected  to  the 
thermocouple  will  give  a  deflection  which  is  a  function  of  the  temperature 
difference  between  the  hot  and  cold  junctions.  Thermocouples  con- 
nected in  series  are  called  thermopiles.  Thermocouples  for  the  measure- 
ment of  high  temperatures  are  calibrated  with  the  aid  of  the  known 
melting  points  of  pure  metals. 

Resistance  thermometers  are  suitable  for  temperature  measurements  up 
to  1800  F.  These  thermometers  depend  for  their  operation  on  the  change 
of  resistance  with  temperature  of  a  platinum,  nickel,  or  copper  wire  coil, 
and  they  are  calibrated  in  the  same  way  as  thermocouples. 

»See  A.S.H.V.E.  research  paper  entitled  Study  of  the  Application  of  Thermocouples  to  the  Measurement 
of  Wall  Surface  Temperatures,  by  A.  P,  Jtratz  and  E.  L,  Broderick  (A.S.H.V.E,  TRANSACTIONS,  Vol.  38, 
1932). 

676 


CHAPTER  40 — TEST  METHODS  AND  INSTRUMENTS 


For  temperatures  above  500  F  various  types  of  pyrometers  are  employed. 
The  mercurial  pyrometer  is  a  thermometer  with  an  inert  gas,  such  as 
nitrogen  or  carbon  dioxide,  above  the  mercury  column  to  prevent  the 
mercury  from  boiling.  The  radiation  pyrometer  consists  of  a  thermopile 
upon  which  the  radiation  from  a  hot  source  is  focused  by  a  concave  mirror. 
A  sensitive  galvanometer  with  a  calibrated  temperature  scale  indicates 
the  thermo-electromotive  force  created  by  the  heat  on  the  thermopile. 
The  optical  pyrometer  measures  radiant  energy  by  comparing  the  intensity 
of  a  narrow  spectral  band,  usually  red  light  emitted  by  the  object,  with 
that  emitted  by  a  standard  light  source  (electric  lamp) .  Thermo-electric 
pyrometers  operate  on  the  same  principle  as  thermocouples.  When 
measuring  high  temperatures,  it  is  customary  to  hold  the  cold  junction  at 
room  temperature  and  this  may  cause  some  error  if  the  room  temperature 
is  above  or  below  the  calibration  point.  For  extremely  precise  tempera- 
ture measurements,  the  cold  junction  is  usually  immersed  in  melting  ice 
to  fix  the  cold  junction  temperature.  Various  forms  of  hand-operated  and 
automatic  cold  junction  temperature  compensators  are  also  available. 

In  the  measuring  of  room  temperatures  care  must  be  exercised  to  pre- 
vent the  results  from  being  affected  by  the  body  heat  of  the  observer, 
by  drafts  from  doors,  windows  and  other  openings,  or  by  radiant  heat 
from  some  local  source  such  as  a  radiator  or  wall.  All  thermometers 
should  be  mercury  thermometers  with  engraved  stems.  The  total  gradua- 
tions of  the  thermometers  should  be  from  20  to  120  F,  in  one  degree 
graduations.  No  ten,  degrees  should  occupy  a  space  of  less  than  one-half 
inch.  The  accuracy  throughout  the  whole  scale  must  be  within  one-half 
degree.  The  operator  should  take  hold  of  the  top  and  no  part  of  the  body, 
including  the  hand,  should  be  nearer  than  10  in.  to  the  bulb.  The 
thermometer  should  not  be  closer  than  5  ft  to  any  door,  window,  or  other 
opening;  should  not  be  closer  than  12  in.  to  any  wall;  and  should  be 
between  3  and  5  ft  from  the  floor.  A  sling  instrument  should  be  used  for 
extreme  accuracy.  Thermocouples  or  resistance  thermometers  may  also 
be  used  for  room  temperature  measurements,  an  advantage  being  that  the 
operator  can  read  temperatures  from  outside  the  room  if  desired,  and  thus 
eliminate  the  errors  which  might  be  caused  by  his  presence  close  to  the 
temperature  measuring  device. 

For  measuring  duct  temperatures  a  duct  thermometer  should  be  used, 
with  the  bulb  extending  into  the  duct  at  least  6  in.  When  the  thermo- 
meter is  to  be  permanently  located  in  the  duct,  a  pipe  flange  or  nipple 
should  be  used  to  receive  the  threaded  portion  of  the  thermometer  stem. 
When  the  thermometer  is  not  to  be  permanently  located,  a  cork  or  rubber 
stopper  may  be  placed  around  the  stem  to  prevent  errors  from  air  leakage. 
Readings  should  be  taken  at  various  locations  in  a  duct  so  due  con- 
sideration may  be  given  to  temperature  stratification.  Other  forms  of 
temperature  measuring  devices  may  be  used,  but  the  active  part  must  be 
at  least  6  inches  from  the  duct  wall. 

Recording  thermometers  may  be  used  for  testing,  and  for  making 
continuous  records  of  operation.  Care  should  be  taken,  however,  to 
insure  that  time  lag  due  to  heavy  measuring  elements  is  kept  to  a  mini- 
mum, so  that  the  recorders  will  properly  follow  temperature  fluctuations. 
Thermocouples  made  of  fine  wire  will  show  less  time  lag  than  will  many 
mercury  bulb  thermometers. 

677 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

MEASUREMENT  OF  AIR  MOVEMENT 

The  quantity,  velocity  and  pressure  of  air  moved  by  a  fan  or  flowing 
through  a  duct  or  grille  may  be  determined  by  various  methods.  The 
instruments  in  common  use  are  the  Pitot  tube,  anemometer,  direct 
reading  velocity  meter,  and  Kata-thermometer,  the  latter  being  suitable 
for  low  air  velocities  and  being  commonly  used  for  measurements  at 
points  where  the  air  is  not  confined  in  a  duct.  The  use  of  calibrated 
nozzles,  orifice  plates,  and  Venturi  meters  are  recognized  methods,  which, 
however,  have  little  application  in  connection  with  ventilation  practice. 

Pitot  Tube 

This  usually  consists  of  two  tubes,  one  within  the  other,  which  when 
properly  held  in  the  air  stream  will  register  the  total  or  impact  pressure 
and  the  static  pressure,  respectively.  If  these  tubes  are  connected  to 
opposite  sides  of  a  water  column,  or  other  type  of  manometer,  the  recorded 
pressure  will  be  the  differential  or  velocity  pressure.  Volume  measure- 
ments may  thus  be  made  in  a  duct  of  known  area.  Pitot  tube  measure- 
ments are  preferably  used  for  air  velocities  exceeding  20  fps.  Volumetric 
determinations  from  Pitot  tube  readings  should  take  into  account  the 
barometric  pressure  and  the  temperature  and  humidity  of  the  air  measured. 

In  general  no  accurate  velocity  pressure  readings  can  be  taken  when  the 
flow  of  air  in  ducts  is  turbulent.  To  insure  accuracy  a  straight  section  of 
duct  from  5  to  10  times  its  own  diameter  is  desirable  in  order  to  straighten 
out  the  air  currents.  If  it  is  necessary  to  take  Pitot  tube  readings  in 
shorter  sections  of  straight  duct,  the  results  must  be  considered  subject 
to  some  doubt  and  checked  accordingly.  For  accurate  work  it  is  neces- 
sary to  make  a  traverse  of  the  duct,  dividing  its  cross  section  into  a 
number  of  imaginary  equal  areas  and  taking  a  reading  in  the  center  of 
each,  the  average^of  the  velocities  corresponding  to  these  pressures  giving 
the  true  velocity  in  the  duct. 

Anemometer 

This  instrument  is  delicate,  and  requires  frequent  calibration  when 
accuracy  is  desired.  The  vanes  of  the  instrument  should  never  be 
touched  and  it  should  never  be  held  in  air  having  a  velocity  greater  than 
that  for  which  it  is  calibrated.  Readings  taken  directly  in  a  fan  inlet  or 
discharge  are  likely  to  harm  the  instrument  because  of  excessive  velocities. 
In  duct  measurements  the  same  procedure  is  followed  as  for  the  Pitot 
tube.  The  anemometer  usually  reads  directly  in  linear  feet.  To  obtain 
the  velocity  in  feet  per  minute,  the  reading  must  be  divided  by  the 
elapsed  time  in  minutes. 

The  following  procedure  for  obtaining  anemometer  readings  is  based 
on  research  conducted  at  Armour  Institute  of  Technology  in  co5peration 
with  the  A.S.H.V.E.  Research  Laboratory2. 

Supply  Grilles.  The  surface  of  the  grille  should  be  marked  off  into  a 
number  of  equal  areas  approximately  6  in.  square.  A  4-in.  anemometer 

'Measurement  of  Flow  of  Air  through  Registers  and  Grilles,  by  I/.  E.  Davies  (A.S,H.V»B,  TltANSACTtOHS^ 
Vol.  36,  1930,  VoL  37,  1931,  and  A.S.H.V.E.  Journal  Section,  Heating,  Piping  **&  Air  Conditioning,  Sep- 
tember, 1933).  »  «.  K  *  r 

678 


CHAPTER  40 — TEST  METHODS  AND  INSTRUMENTS 


should  be  used  and  should  be  held  at  the  center  of  each  section  in  contact 
with  the  grille  (or  as  close  as  possible)  for  a  period  of  time  sufficient  to 
insure  an  average  reading.  In  the  case  of  supply  grilles,  the  instrument 
should  always  be  held  with  the  dial  facing  the  operator.  The  average  of 
the  corrected  readings  should  then  be  used  in  the  following  formula  to 
obtain  the  flow  in  cubic  feet  per  minute: 


where 

V  —  average  of  corrected  anemometer  readings,  feet  per  minute. 
A  =  gross  area  of  grille,  square  feet. 
a  =  net  free  area  of  grille,  square  feet. 
p  —  percentage  of  free  area  of  grille  expressed  as  a  decimal. 
C  «=  a  coefficient  that  varies  with  the  velocity  from  grille  and  may  vary  slightly 
with  type  of  grille.    For  average  use,  with  supply  grilles,  C  can  be  taken 
as  0.97  at  velocities  from  150  to  600  fpm,  and  as  1.00  at  higher  velocities. 

Particular  care  should  be  exercised  in  the  case  of  long,  narrow  grilles. 
The  nature  of  the  approach  sometimes  results  in  there  being  a  narrow 
strip  along  the  top  or  bottom  of  the  grille  through  which  no  air  will  be 
flowing.  This  may  be  detected  by  holding  the  anemometer  completely 
out  of  the  air  stream  and  then  moving  it  slowly  inward  over  the  grille  until 
the  vanes  just  start  to  move.  The  distance  which  the  vanes  extend  over 
the  grille  opening  at  this  moment  will  indicate  the  width  of  the  dead  strip. 
Only  the  remaining  portion  of  the  grille  should  be  considered  in  making 
the  calculations  for  gross  and  free  area. 

Exhaust  Grilles.  The  surface  of  the  grille  should  be  marked  off  and 
readings  taken  in  the  same  manner  as  with  supply  grilles,  except  that  the 
instrument  should  be  held  with  the  dial  facing  the  grille,  and  in  contact 
with  it.  The  traverse  should  be  taken  at  a  uniform  rate,  allowing  suf- 
ficient time  in  each  space  to  minimize  the  percentage  of  error.  In  the  case 
of  exhaust  grilles  it  is  found  that  the  formula 

tfm  «  KVA  (2) 

in  which 

V  «  average  indicated  velocity  obtained  by  the  anemometer  traverse. 
4  =  gross  area  of  grille,  square  feet. 

j£  =  coefficient  determined  by  experiment.    For  average  use,  with  exhaust  grilles, 
K  may  be  taken  as  0.8  for  all  usual  velocities. 

This  formula  is  of  advantage,  especially  with  ornamental  grilles,  in 
that  the  free  area  need  not  be  measured. 

The  flow  of  air  through  registers  and  grilles  is  of  considerable  impor- 
tance, being  frequently  the  only  convenient  method  of  measuring  the 
volume  of  supply  air  to  a  room.  While  duct  measurements,  if  available, 
are  more  dependable,  grille  measurements  provide  a  fairly  accurate 
method,  if  care  is  taken  in  the  technique  of  using  the  anemometer. 

Kata-Thermometer 

The  Kata-thermometer  can  be  used  to  determine  air  velocities  pro- 
vided the  walls  and  surrounding  objects  are  at  or  near  the  room  tern- 

679 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

perature.  Especially  at  low  velocities  it  constitutes  a  useful  instrument 
for  readily  detecting  drafts. 

The  instrument  is  essentially  an  alcohol  thermometer  with  a  bulb 
approximately  5^  in.  in  diameter  and  J^  in.  long  with  a  stem  8  in.  long 
reading  from  100  F  to  95  F,  graduated  to  tenths  of  a  degree.  To  take 
readings  the  bulb  is  heated  in  water  until  the  alcohol  expands  and  rises 
into  a  top  reservoir.  The  time  in  seconds  required  for  the  liquid  to 
fall  from  100  F  to  95  F  is  recorded  with  a  stop  watch  and  this  time  is  a 
measure  of  the  rate  of  cooling. 

The  dry  Kata  loses  its  heat  by  radiation  and  by  convection  so  for 
constant  velocities  the  time  of  cooling  is  a  function  of  the  dry-bulb  tem- 
perature of  the  surrounding  air.  The  wet  Kata,  which  has  a  cloth  covering 
fitted  snugly  around  its  bulb,  loses  heat  by  radiation,  convection,  and 
evaporation,  and  for  constant  velocities  its  rate  of  cooling  is  a  function  of 
the  wet-bulb  temperature  of  the  air  irrespective  of  the  dry-bulb  tem- 
perature or  relative  humidity.  It  does  not  follow,  however,  that  the 
difference  in  rate  of  cooling  of  the  dry  and  the  wet  Kata  is  caused  by 
evaporation.  A  change  in  the  wet-bulb  temperature  produces  a  change  in 
the  surface  temperature  of  the  wet  Kata  which  in  turn  affects  the  heat 
lost  by  radiation  and  by  convection. 

Several  precautions  should  be  taken  to  obtain  the  best  results  with  this 
instrument: 

1.  To  obtain  velocity  readings  use  the  dry  Kata  since  the  error  in  timing  is  reduced. 

2.  The  instrument  should  be  heated  and  allowed  to  cool  two  or  three  times  before 
recording  the  final  time  of  cooling.    The  first  reading  is  not  reliable. 

3.  All  traces  of  moisture  must  be  removed  from  the  dry  Kata  before  timing  to  eli- 
minate error  introduced  by  evaporation. 

4.  Use  only  the  formula  applying  to  a  particular  instrument.    Each  Kata  receives  an 
individual  calibration. 

HUMIDITY  MEASUREMENT 

The  sling  psychrometer  is  the  recognized  standard  instrument  for 
determining  humidities.  In  order  to  obtain  accurate  readings  considerable 
skill  is  required  on  the  part  of  the  operator.  The  wicking  and  water  must 
be  clean  and  the  temperature  of  the  water  should  be  slightly  above  the 
wet-bulb  temperature  of  the  surrounding  air.  The  psychrometer  should 
be  swung  rapidly  and  several  and  frequent  observations  should  be  made 
to  see  that  the  wet-bulb  temperature  has  become  stationary  before  the 
final  reading  is  noted.  Care  should  be  taken  that  the  wet-bulb  has 
reached  a  minimum  temperature,  but  the  wick  must  still  be  moist. 
Standard  psychrometric  tables  should  be  used. 

In  making  wet-bulb  measurements  below  32  F  the  same  procedure  is 
followed  as  above  32  F.  The  water  is  liquid  at  the  start,  but  as  the  sling 
is  operated  it  will  freeze  rapidly  enough  so  that  in  quickly  giving  up  the 
latent  heat  of  fusion,  the  indicated  wet-bulb  temperature  may  drop 
below  the  actual  wet-bulb  temperature.  After  the  liquid  on  the  bulb  has 
become  thoroughly  frozen  the  wet-bulb  temperature  will  rise  to  normal. 
A  very  thin  film  of  ice  is  more  desirable  than  a  thick  film.  Care  must  be 
taken  to  read  the  temperatures  in  the  region  below  32  F  accurately 
because  the  spread  between  the  wet-  and  dry-bulb  is  small, 

680 


CHAPTER  40 — TEST  METHODS  AND  INSTRUMENTS 


In  taking  humidity  readings  in  ducts  it  is  usually  impracticable  to  use 
a  sling  psychrometer.  For  this  work  the  stationary  hygrodeik  arranged 
for  bolting  on  to  the  side  of  the  duct,  with  two  bulbs  extending  into  the 
duct,  will  be  found  very  convenient.  Owing  to  the  velocity  of  the  air 
passing  over  the  bulbs  within  the  duct  an  accurate  reading  will  be  secured, 
corresponding  to  that  given  by  the  sling  psychrometer. 

Various  forms  of  humidity  recorders  are  available,  some  merely  re- 
cording wet-  and  dry-bulb  temperatures,  and  others  recording  relative 
humidity  directly.  Any  form  of  wet-  and  dry-bulb  device  must  have 
sufficient  air  velocity  over  the  thermometer  bulbs  to  insure  accurate 
readings ;  this  velocity  should  be  secured  by  a  fan  if  the  air  is  not  itself  in 
motion,  as  in  a  duct.  For  extremely  low  humidities,  or  for  humidity 
measurements  above  212  F,  a  thermal  conductivity  method  is  available3. 

CARBON  DIOXIDE  DETERMINATION4 

At  ordinary  concentrations  carbon  dioxide  is  not  harmful.  The  amount 
of  carbon  dioxide  in  the  air  is  a  convenient  index  of  the  rate  of  air  supply, 
and  of  the  distribution  of  the  air  within  rooms.  Unequal  carbon  dioxide 
concentrations  in  parts  of  a  room  indicate  improper  air  distribution. 

The  Petterson-Palmquist  apparatus  has  been  generally  accepted  as  the 
standard  device  for  the  determination  of  carbon  dioxide  in  air  investiga- 
tions. The  principle  involved  is  the  measurement  of  a  given  volume  of 
air,  the  absorption  of  the  contained  carbon  dioxide  in  a  caustic  potash 
solution,  and  the  remeasurement  of  the  volume  of  air  at  the  original 
pressure  in  a  finely  graduated  capillary  tube,  the  difference  in  volume 
representing  the  absorbed  carbon  dioxide.  (See  Report  of  Committee  on 
Standard  Methods  for  Examination  of  Air,  American  Public  Health  Asso- 
ciation, Vol.  7,  No.  1;  American  Journal  of  Public  Health,  Jan.,  1917.) 

Where  field  conditions  are  such  that  this  apparatus  may  not  be  con- 
veniently used,  as  in  street  cars,  air  samples  may  be  collected  in  clean 
bottles  having  mercury-sealed  rubber  stoppers,  and  these  may  be  sub- 
jected to  laboratory  analysis. 

DUST  DETERMINATION 

Many  laboratory  methods  have  been  developed  to  measure  the  dust  in 
the  air.  These  involve  the  collection  of  dust  on  sticky  plates,  on  filter 
paper,  in  water,  on  porous  crucibles,  or  by  electric  precipitation,  and  the 
subsequent  determination  of  the  amount  of  dust  by  microscopic  counting, 
weighing,  or  titration.  While  there  is  no  standard  method,  the  Hill 
dust  counter,  using  a  microscope,  the  impinger6,  using  chemical  changes 
in  water,  and  the  Lewis  sampling  tube6,  involving  the  analytical  weighing 
of  a  porous  crucible,  are  accepted.  All  test  results  should  be  accompanied 
by  the  name  of  the  instrument  used  as  great  variation  in  counts  with  the 


'"Gas  Analysis  by  Measurement  of  Thermal  Conductivity,"  H.  A.  Daynes,  Cambridge  Press,  1933. 

*See  A.S.H.V.E.  research  paper  entitled  Indices  of  Air  Change  and  Air  Distribution,  by  F.  C.  Houghten 
and  J.  L,  Blackshaw  (A.S,H.V,E.  Journal  Section,  Keating,  Piping  and  Air  Conditioning,  June,  1933). 

'Public  Health  Bulletin,  No.  144,  1925,  U.  S.  Public  Health  Service. 

•Testing  and  Rating  of  Air  Cleaning  Devices  Ueed  for  General  Ventilation  Work,  by  Samuel  R.  Lewis 
(A.S.H.V.E.  Journal  Section^  Heating,  Piping  and  Air  Conditioning,  May,  1933), 

681 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

different  instruments  will  be  obtained.  The  AMERICAN  SOCIETY  OF 
HEATING  AND  VENTILATING  ENGINEERS  has  developed  a  code7  for  the 
testing  and  rating  of  air  cleaning  devices  used  in  general  ventilation  work. 

FLUE  CAS  ANALYSIS 

The  analysis  of  flue  gases  by  chemical  means  is  made  with  the  Orsat 
apparatus.  A  solution  of  KOH  is  used  to  absorb  the  C02.  Free  oxygen  is 
absorbed  by  a  mixture  of  pyrogallic  acid  and  KOH.  The  solution  for 
absorbing  the  CO  is  cuprous  chloride.  The  apparatus  consists  of  a 
burette  surrounded  by  a  water  jacket,  to  receive  and  measure  the  volume 
of  gas.  The  burette  is  connected  by  a  manifold  of  glass  to  pipettes  con- 
taining liquids  for  absorbing  CO^  0%  and  CO. 

Various  forms  of  automatic  indicating  and  recording  gas  analysis 
devices  are  available,  operating  on  either  chemical  or  physical  principles. 
Such  devices  are  convenient  for  plant  operation. 

MEASUREMENT  OF  SMOKE  DENSITY 

Relative  smoke  density  is  usually  measured  by  comparison  with  the 
Ringelmann  Chart  (Fig.  1).  In  making  observations  of  the  smoke  issuing 
from  a  chimney,  four  cards  ruled  like  those  in  Fig.  1,  together  with  a  card 
printed  in  solid  black  and  another  left  entirely  white,  are  placed  in  a 
horizontal  row  and  hung  at  a  point  50  ft  from  the  observer  and  con- 
veniently in  line  with  the  chimney.  At  this  distance,  the  lines  become 
invisible,  and  the  cards  appear  to  be  of  different  shades  of  gray,  ranging 
from  very  light  gray  to  almost  black.  The  observer  glances  from  the 
smoke  coming  from  the  chimney  to  the  cards,  which  are  numbered  from 
0  to  5,  determines  which  card  most  nearly  corresponds  with  the  color  of 
the  smoke,  and  makes  a  record  accordingly,  noting  the  time.  Observa- 
tions are  made  continuously  during  one  minute,  and  the  estimated  average 
density  during  that  minute  recorded.  The  average  of  all  the  records 
made  during  a  boiler  test  is  taken  as  the  average  figure  for  the  smoke 
density  during  the  test,  and  the  entire  record  is  plotted  on  cross-section 
paper  in  order  to  show  how  the  smoke  varied  in  density  from  time  to  time* 

Smoke  Recorders 

Smoke  recorders  are  available  which  give  a  much  more  accurate  in- 
dication of  the  amount  of  smoke  being  produced  than  does  the  Ringel- 
mann Chart.  Although  most  of  these  Instruments  are  in  the  process  of 
development,  they  constitute  a  satisfactory  tool  in  the  control  of  smoke 
emission.  They  all  depend  upon  projecting  a  beam  of  light  through  the 
smoke  flue  or  through  a  separate  compartment  from  which  a  sample  of  the 
flue  gas  is  drawn  continuously.  The  light  of  the  beam  which  passes 
through  without  being  absorbed  by  the  smoke  is  measured  to  determine 
the  smoke  density.  Most  of  these  instruments  make  use  of  a  photo- 
electric cell  or  a  thermopile  to  measure  the  relative  amount  of  light  which 
has  not  been  absorbed.  Standard  electrical  instruments  serve  for  in- 
dicating or  recording. 


'See  A.S.H.V.E.  Standard  Code  for  Testing  and  Rating  Air  Cleaning  Devices  Used  in  General  Ventila- 
tion Work,  edition  of  July,  1934, 

682 


CHAPTER  40 — TEST  METHODS  AND  INSTRUMENTS 


MEASUREMENT  OF  RATE  OF  HEAT  TRANSMISSION 

The  standard  methods  of  testing  built-up  wall  sections  are  by  means  of 
the  guarded  hot-box*  and  the  guarded  hot-plate*.  The  Nicholls  heat-flow 
meter9  may  be  used  for  testing  actual  walls  of  buildings. 

It  would  be  obviously  impossible  to  determine  the  air-to-air  heat  trans- 
mission coefficients  of  every  type  of  wall  construction  in  use  with  the 
heat-flow  meter,  the  guarded  hot-box  or  the  guarded  hot-plate  on  account 
of  the  great  amount  of  time  involved.  Hence,  the  method  of  computing 
the  coefficients  from  the  fundamental  constants  must  be  resorted  to  in 
most  cases.  The  guarded  hot-plate  is  used  to  determine  the  fundamental 


FlG.  1.      RlNGELMANN  SMOKE  CHART 

constants.  The  heat-flow  meter,  guarded  hot-box  and  guarded  hot-plate 
tests  can  be  used  to  good  advantage  in  checking  the  accuracy  of  the 
computed  values. 

If  the  hot-box  or  hot-plate  methods  are  used,  tests  are  usually  run  under 
still  air  conditions,  which  means  there  is  no  wind  movement  over  the 
surfaces  of  the  wall  during  the  test.  In  the  hot-plate  method  of  test  the 
inside  surface  coefficient  is  eliminated  by  the  plate's  being  in  direct  contact 
with  the  wall.  In  practice,  some  wind  movement  over  the  exterior  surface 
of  the  wall  should  always  be  allowed  for;  hence,  still-air  coefficients  cannot 
be  used  over  the  outside  of  the  building  during  the  heating  season. 
Moreover,  still-air  transmission  coefficients  cannot  be  corrected  to  provide 
for  moving-air  conditions  by  applying  a  single  constant  factor.  Computed 
coefficients  of  transmission  for  various  types  of  construction  are  given 
in  Chapter  5. 

EUPATHEOSCOPE 

The  eupatheoscope  affords  a  means  of  evaluating  the  combined  effect  of 
radiation  and  convection  in  a  given  environment  in  terms  of  a  standard 
environment  and  in  some  terms  related  to  human  comfort.  See  Chapter 
38. 

•See  Standard  Code  for  Heat  Transmission  through  Walls  (A.S.H.V.E.  TRANSACTIONS,  Vol.  34,  1928> 
and  Report  of  the  Committee  on  Heat  Transmission,  National  Research  Council. 

•See  Measuring  Heat  Transmission  in  Building  Structures  and  a  Heat  Transmission  Meter,  by  P.. 
Nlcholls  (A,S,H,VJ&  TRANSACTIONS,  Vol.  30,  192*7, 

683 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


PROBLEMS  IX  PRACTICE 

1  •  The  hand  on  a  pressure  gage  attached  to  a  steam  line  indicates  a  pressure 
of  15  Ib  per  square  inch  and  the  barometric  pressure  is  14.7  Ib  per  square  inch. 
What  is  the  absolute  pressure,  in  pounds  per  square  inch,  being  exerted  by 
the  steam? 

The  absolute  pressure  exerted  by  the  steam  in  the  pipe  is  equal  to  the  pressure  indicated 
by  the  gage  plus  that  exerted  by  the  atmosphere. 

Total  pressure  =  15  -f  14.7  ==  29.7  Ib  per  square  inch. 

2  •  Outline  the  procedure  to  be  followed  in  taking  room  temperatures. 

In  taking  room  temperatures,  a  standard  mercury  thermometer  should  be  used,  with 
care  taken  that  no  part  of  the  observer's  body  is  nearer  than  10  in.  to  the  thermometer 
bulb.  The  thermometer  should  be  held  at  least  5  ft  away  from  any  window,  door  or 
opening;  it  should  be  at  least  12  in.  away  from  any  wall,  and  should  be  between  3  and  5 
ft  from  the  floor. 

3  •  What  advantages  other  than  its  sensitiveness,  has  the  U-tube  draft  gage  or 
manometer  for  measurement  of  low  pressures? 

Inherent  accuracy  without  calibration  and  low  cost  of  the  essential  parts,  which  are 
glass  tubing  and  an  ordinary  scale. 

4  •  Are  thermocouples  as  accurate  as  mercury  thermometers? 

Within  the  range  which  can  be  measured  with  both  instruments  (below  1000  F)  either 
one  may  be  made  as  sensitive  as  the  service  requires.  The  accuracy  of  a  thermocouple 
temperature  measurement  depends  chiefly  on : 

1.  An  accurate  calibration  of  the  wire. 

2.  The  sensitiveness  of  the  electrical  instrument. 

3.  Accurate  cold-junction  control. 

4.  Proper  placement  of  the  sensitive  junction. 

5  •  Is  room  temperature  accurately  measured  by  the  ordinary  wall  thermo- 
meter? 

No,  Wall  thermometer  measurements  may  be  several  degrees  in  error  as  compared  with 
an  observation  properly  made  in  the  zone  of  occupancy, 

6  •  When  an  anemometer  is  used  for  measuring  the  air  discharged  from  a 
grille  or  register,  does  it  read  the  velocity  through  the  gross  face  area  or  the 
velocity  through  the  net  free  area? 

Neither.  If  either  of  these  velocities  is  required,  it  should  be  calculated  by  means  of 
Equation  1. 

7  •  Do  common  errors  made  in  humidity  determination  produce  a  result  that 
is  too  high  or  too  low? 

A  higher  relative  humidity  than  the  true  value  is  likely  to  be  found,  either  because  there 
is  insufficient  velocity  over  the  wet-bulb  or  because  the  reading  is  not  taken  at  the  right 
time. 

8  •  What  is  the  purpose  of  the  carbon  dioxide  determination? 

It  is  an  index  of  the  adequacy  of  fresh  air  supply  and  also  an  indicator  of  air  distribution. 


684 


Chapter  41 

TERMINOLOGY 

Glossary  of  Physical  and  Heating  and  Ventilating   Terms   Used 

in    the    Text,    Standard    Abbreviations,    Conversion    Equations, 

Drafting  Symbols,  A.S.H.V.E.   Codes 

Absolute  Humidity:    See  Humidity. 

Absolute  Pressure:  The  sum,  at  any  particular  time,  of  the  gage 
pressure  and  the  atmospheric  pressure. 

Absolute  Temperature:  The  temperature  of  a  substance  measured 
above  absolute  zero. 

Absolute  Zero:  The  temperature  (  —  459.6  F)  at  which  the  molecular 
motion  of  a  substance  theoretically  ceases.  This  is  the  temperature  at 
which  the  substance  theoretically  contains  no  heat  energy. 

Acceleration:  The  rate  of  change  of  velocity.  In  the  fps  system 
this  is  expressed  in  units  of  one  foot  per  second  per  second. 


Acceleration  Due  to  Gravity  :  The  rate  of  gain  in  velocity  of  a  freely 
falling  body.  In  the  fps  system  this  is  32.174  feet  per  second  per  second. 

Adiabatic:  An  adjective  pertaining  to  or  designating  variations  in 
volume  or  pressure  not  accompanied  by  gain  or  loss  of  heat.  When  a 
substance  undergoes  adiabatic  expansion,  since  it  does  not  receive  heat 
from  without,  the  work  which  it  does  is  at  the  expense  of  its  internal 
energy,  and  therefore  its  temperature  falls;  similarly,  when  it  is  adia- 
batically  compressed  its  temperature  rises. 

Adsorption:  The  adhesion  of  the  molecules  of  gases  or  dissolved  sub- 
stances to  the  surfaces  of  solid  bodies,  resulting  in  a  concentration  of  the 
gas  or  solution  at  the  place  of  contact, 

Air  Cleaner:  A  device  designed  for  the  purpose  of  removing  air-borne 
impurities  such  as  dusts,  fumes,  and  smokes.  (Air  cleaners  include  air 
washers  and  air  filters,) 

Air  Conditioning:  The  simultaneous  control  of  all  or  at  least  the  first 
three  of  those  factors  affecting  both  the  physical  and  chemical  conditions 
of  the  atmosphere  within  any  structure.  These  factors  include  tempera- 
ture, humidity,  motion,  distribution,  dust,  bacteria,  odors,  toxic  gases, 
and  ionization,  most  of  which  affect  in  greater  or  lesser  degree  human 
health  or  comfort. 

Air  Infiltration  :  The  inleakage  of  air  through  cracks  and  crevices, 
and  through  doors,  windows  and  other  openings,  caused  by  wind  pressure 
or  temperature  difference. 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

Air  Washer:  An  enclosure  in  which  air  is  forced  through  a  spray  of 
water  in  order  to  cleanse,  humidify,  or  dehumidify  the  air. 

Anemometer:  An  instrument  for  measuring  the  velocity  of  moving 
air, 

Atmospheric  Pressure:  The  pressure  exerted  by  the  atmosphere  in 
all  directions,  as  indicated  by  a  barometer.  Standard  atmospheric  pressure 
is  considered  to  be  14.7  Ib  per  square  inch,  which  is  equivalent  to  29.92  in. 
of  mercury. 

Baffle:    A  plate  or  wall  for  deflecting  gases  or  fluids. 

Blast:  This  word  was  formerly  used  to  denote  forced  air  circulation, 
particularly  in  connection  with  central  fan  systems  using  steam  or  hot 
water  as  the  heating  medium.  As  applied  in  this  sense,  the  word  blast 
is  now  obsolete. 

Boiler:  A  closed  vessel  in  which  steam  is  generated  or  in  which  water  is 
heated. 

Boiler  Heating  Surface:  That  portion  of  the  surface  of  the  heat- 
transfer  apparatus  in  contact  with  the  fluid  being  heated  on  one  side  and 
the  gas  or  refractory  being  cooled  on  the  other,  in  which  the  fluid  being 
heated  forms  part  of  the  circulating  system  ;  this  surface  shall  be  measured 
on  the  side  receiving  heat.  This  includes  the  boiler,  water  walls,  water 
screens,  and  water  floor.  (A.S.M.E.  Power  Test  Codes,  Series  1929.) 

Boiler  Horsepower:  The  equivalent  evaporation  of  34.5  Ib  of  water 
per  hour  from  and  at  212  F.  This  is  equal  to  a  heat  output  of  970.2  X 
34.5  =  33,471.9  Btu  per  hour. 

British  Thermal  Unit:    The  mean  British  thermal  unit  is  -      of  the 


heat  required  to  raise  the  temperature  of  1  Ib  of  water  from  32  F  to  212  F. 
It  is  substantially  equal  to  the  quantity  of  heat  required  to  raise  1  Ib  of 

water  from  63  F  to  64  F.    One  Btu  =  ^-5"  kwhn 

By-pass  :  A  pipe  or  duct,  usually  controlled  by  valve  or  damper,  for 
short-circuiting  fluid  flow. 

Calorie:    The  mean  calorie  is  -r-nn  °f  ^e  heat  required  to  raise  the 

J.UU 

temperature  of  1  gram  of  water  from  Zero  C  to  100  C.  It  is  substantially 
equal  to  the  quantity  of  heat  required  to  raise  one  gram  of  water  from 
14.5  C  to  15.5  C. 

Central  Fan  System:  A  mechanical  indirect  system  of  heating, 
ventilating,  or  air  conditioning,  in  which  the  air  is  treated  or  handled  by 
equipment  located  outside  the  rooms  served,  usually  at  a  central  location, 
and  is  conveyed  to  and  from  the  rooms  by  means  of  a  fan  and  a  system  of 
distribution  ducts.  See  Chapters  9  and  22. 

Chimney  Effect:  The  tendency  in  a  duct  or  other  vertical  air  passage 
for  air  to  rise  when  heated,  owing  to  its  decrease  in  density. 

Coefficient  of  Transmission:  The  amount  of  heat  (Btu)  transmitted 
from  air  to  air  in  one  hour  per  square  foot  of  the  wall,  floor,  roof  or  ceiling 
for  a  difference  in  temperature  of  1  F  between  the  air  on  the  inside  and  that 
on  the  outside  of  the  wall,  floor,  roof  or  ceiling. 

686 


CHAPTER  41 — TERMINOLOGY 


Column  Radiator:  A  type  of  direct  radiator.  This  radiator  has  not 
been  listed  by  manufacturers  since  1926. 

Comfort  Line:  The  effective  temperature  at  which  the  largest  per- 
centage of  adults  feel  comfortable. 

Comfort  Zone'  (Average):  The  range  of  effective  temperatures  over 
which  the  majority  (50  per  cent  or  more)  of  adults  feel  comfortable. 
Comfort  Zone  (Extreme):  The  range  of  effective  temperatures  over  which 
one  or  more  adults  feel  comfortable.  (See  Chapter  2.) 

Concealed  Radiator :   See  Convector. 

Conductance:  The  amount  of  heat  (Btu)  transmitted  from  surface 
to  surface  in  one  hour  through  one  square  foot  of  a  material  or  construc- 
tion, whatever  its  thickness,  when  the  temperature  difference  is  1  F 
between  the  two  surfaces. 

Conduction;  The  transmission  of  heat  through  and  by  means  of 
matter  unaccompanied  by  any  obvious  motion  of  the  matter. 

Conductivity:  The  amount  of  heat  (Btu)  transmitted  in  one  hour 
through  one  square  foot  of  a  homogeneous  material  1  in.  thick  for  a 
difference  in  temperature  of  1  F  between  the  two  surfaces  of  the  material. 

Conductor  (heat):  A  material  capable  of  readily  conducting  heat. 
The  opposite  of  an  insulator  or  insulation. 

Constant  Relative  Humidity  Line:  Any  line  on  the  psychrometric 
chart  representing  a  series  of  conditions  which  may  be  evaluated  by  one 
percentage  of  relative  humidity;  there  are  also  constant  dry-bulb  lines, 
wet-bulb  lines,  effective  temperature  lines,  vapor  pressure  lines,  and 
lines  showing  other  physical  properties  of  air  mixed  with  water  vapor. 

Control:  Any  manual  or  automatic  device  for  the  regulation  of  a 
machine  to  keep  it  at  normal  operation.  If  .automatic,  it  is  considered 
that  the  device  is  motivated  by  variations  in  temperature,  pressure, 
time,  light,  or  other  influences. 

Convection:  The  transmission  of  heat  by  the  circulation  of  a  liquid 
or  a  gas  such  as  air.  Convection  may  be  natural  or  forced. 

Convector:  A  concealed  radiator.  A  heating  unit  and  an  enclosure 
or  shield  located  either  within,  adjacent  to,  or  exterior -to  the  room  or 
space  to  be  heated,  but  transferring  heat  to  the  room  or  space  mainly  by 
the  process  of  convection.  If  the  heating  unit  is  located  exterior  to  the 
room  or  space  to  be  heated,  the  heat  is  transferred  through  one  or  more 
ducts  or  pipes;  see  Chapter  30. 

Corrosive:  Having  the  power  to  wear  away  or  gradually  change  the 
texture  or  substance  of  a  material. 

Decibel:  The  standard  unit  for  noise  or  sound  intensity.  One  decibel 
is  equal  to  ten  times  the  logarithm  to  the  base  e  of  the  ratio  of  the  sound 
intensities. 

Degree-Day:  "A  unit,  based  upon  temperature  difference  and  time, 
used  in  specifying  the  nominal  heating  load  in  winter.  For  any  one  day 
there  exist  as  many  degree-days  as  there  are  degrees  Fahrenheit  dif- 
ference in  temperature  between  the  average  outside  air  temperature, 
taken  over  a  24-hour  period,  and  a  temperature  of  65  F. 

687 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

Dehumidify :  To  remove  water  vapor  from  the  atmosphere;  to 
remove  water  vapor  or  moisture  from  any  material. 

Density:    The  weight  of  a  unit  volume,  expressed  in  pounds  per  cubic 

W 
foot,    p  (rho)  =  -~r. 

Dew-Point  Temperature :  The  temperature  corresponding  to  satura- 
tion (100  per  cent  relative  humidity)  for  a  given  moisture  content. 

Diffuser:  A  vaned  device  placed  at  an  air  supply  opening  to  direct  the 
air  flow. 

Direct-Indirect  Heating  Unit:  A  heating  unit  located  in  the  room 
or  space  to  be  heated  and  partially  enclosed,  the  enclosed  portion  being 
used  to  heat  air  which  enters  from  outside  the  room. 

Direct  Radiator :  Same  as  radiator. 

Direct-Return  System  (Hot  water):  A  hot  water  system  in  which  the 
water,  after  it  has  passed  through  a  heating  unit,  is  returned  to  the  boiler 
along  a  direct  path  so  that  the  total  distance  traveled  by  the  water  is  the 
shortest  feasible,  and  so  that  there  are  considerable  differences  in  the 
lengths  of  the  several  circuits  composing  the  system. 

Down-Feed  One-Pipe  Riser  (Steam):  A  pipe  which  carries  steam 
downward  to  the  heating  units  and  into  which  the  condensation  from  the 
heating  units  drains. 

Down-Feed  System  (Steam):  A  steam  heating  system  in  which  the 
supply  mains  are  above  the  level  of  the  heating  units  which  they  serve. 

Draft  Head  (Side  Outlet  Enclosure) :  The  height  of  a  gravity  convector 
between  the  bottom  of  the  heating  unit  and  the  bottom  of  the  air  outlet 
opening. 

Draft  Head  (Top  Outlet  Enclosure):  The  height  of  a  gravity  convector 
between  the  bottom  of  the  heating  unit  and  the  top  of  the  enclosure. 

Dry  Air:  Air  with  which  no  water  vapor  is  mixed.  This  term  is  used 
comparatively,  since  in  nature  there  is  always  some  water  vapor  included 
in  air,  and  such  water  vapor,  being  a  gas,  is  dry. 

Dry -Bulb  Temperature:  The  temperature  of  the  air  indicated  by 
any  type  of  thermometer  not  affected  by  the  water  vapor  content  or 
relative  humidity  of  the  air. 

Dry  Return:  A  return  pipe  in  a  steam  heating  system  which  carries 
both  water  of  condensation  and  air.  See  wet  return. 

Dust:  Solid  material  in  a  finely  divided  state,  the  particles  of  which 
are  large  and  heavy  enough  to  fall  with  increasing  velocity,  due  to  gravity 
in  still  air.  For  instance,  particles  of  fine  sand  or  grit,  the  average 
diameter  of  which  is  approximately  0.01  centimeter,  such  as  are  blown 
on  a  windy  day,  may  be  called  dust. 

Dynamic  Head  or  Pressure:  The  total  or  impact  pressure.  This  is 
the  sum  of  the  radial  pressure  and  the  velocity  pressure  at  the  point  of 
measurement. 

Effective  Temperature:  An  arbitrary  index  of  the  degree  of  warmth 
or  cold  felt  by  the  human  body  in  response  to  temperature,  humidity, 
and  movement  of  the  air.  Effective  temperature  is  a  composite  index 

688 


CHAPTER  41 — TERMINOLOGY 


which  combines  the  readings  of  temperature,  humidity,  and  air  motion 
into  a  single  value.  The  numerical  value  of  the  effective  temperature 
scale  has  been  fixed  by  the  temperature  of  saturated  air  which  induces  an 
identical  sensation  of  warmth. 

Enthalpy:    Total  heat  or  thermal  potential. 

Entropy:  The  logarithmic  probability  of  a  state.  It  is  the  integra- 
tion between  two  absolute  temperatures  of  the  quotient  of  the  quantity 
of  heat  divided  by  the  absolute  temperature  at  the  condition  at  which  the 
temperature  is  taken.  It  is,  therefore,  a  numeric  which  explains  a  dif- 
ference in  conditions  between  two  points  in  a  heat  cycle. 

Entropy,  which  can  vary  with  temperature,  volume,  or  pressure,  is 
constant  during  adiabatic  expansion  in  a  reversible  cycle  or  during 
isentropic  expansion  in  an  irreversible  cycle.  Entropy  is  a  function  of  the 
unavailable  energy  in  any  system. 

Equivalent  Evaporation:  The  amount  of  water  a  boiler  would 
evaporate,  in  pounds  per  hour,  if  it  received  feed  water  at  212  F  and 
vaporized  it  at  the  same  temperature  and  atmospheric  pressure. 

Estimated  Design  Load:  The  load,  stated  in  Btu  per  hour  or  equiv- 
alent direct  radiation,  as  estimated  by  the  purchaser  for  the  conditions  of 
inside  and  outside  temperature  for  which  the  amount  of  installed  radiation 
was  determined.  It  is  the  sum  of  the  heat  emission  of  the  radiation  to  be 
actually  installed  plus  the  allowance  for  the  heat  loss  of  the  connecting 
piping  plus  the  heat  requirement  for  any  apparatus  requiring  heat  con- 
nected with  the  system.  (A.S.H.V.E.  Standard  Code  for  Rating  Steam 
Heating  Solid  Fuel  Hand-Fired  Boilers— edition  of  April  1932.) 

Estimated  Maximum  Load:  Construed  to  mean  the  load  stated  in 
Btu  per  hour  or  equivalent  direct  radiation  that  has  been  estimated  by 
the  purchaser  to  be  the  greatest  or  maximum  load  that  the  boiler  will  be 
called  upon  to  carry.  (A.S.H.V.E.  Standard  Code  for  Rating  Steam 
Heating  Solid  Fuel  Hand-Fired  Boilers— edition  of  April  1932.) 

Extended  Heating  Surface:  See  Heating  Surface. 

Extended  Surface  Heating  Unit:  A  heating  unit  having  a  relatively 
large  amount  of  extended  surface  which  may  be  integral  with  the  core 
containing  the  heating  medium  or  assembled  over  such  a  core,  making 
good  thermal  contact  by  pressure  or  by  being  soldered  to  the  core  or  by 
both  pressure  and  soldering.  An  extended  surface  heating  unit  is  usually 
placed  within  an  enclosure  and  therefore  functions  as  a  convector. 

Fan  Furnace  System:  See  Warm  Air  Heating  System. 

Force:    The  action  on  a  body  which  tends  to  change  its  relative  con- 

WV 
dition  as  to  rest  or  motion.    F  =  — -. 

Fumes:  Particles  of  solid  matter  resulting  from  such  chemical  pro- 
cesses as  combustion,  explosion,  and  distillation,  ranging  from  0.1  to  1.0 
micron  in  size. 

Furnace:  That  part  of  a  boiler  or  warm  air  heating  plant  in  which 
combustion  takes  place.  Also,  a  fire-pot. 

Furnace  Volume  (Mai):  The  total  furnace  volume  for  horizontal- 
return  tubular  boilers  and  water-tube  boilers  is  the  cubical  contents  of  the 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

furnace  between  the  grate  and  the  first  plane  of  entry  into  or  between 
tubes.  It  therefore  includes  the  volume  behind  the  bridge  wall  as^in 
ordinary  horizontal -return  tubular  boiler  settings,  unless  manifestly  in- 
effective (i.e.,  no  gas  flow  taking  place  through  it),  as  in  the  case  of  waste- 
heat  boilers  with  auxiliary  coal  furnaces,  where  one  part  of  the  furnace  is 
out  of  action  when  the  other  is  being  used.  For  Scotch  or  other  internally 
fired  boilers  it  is  the  cubical  contents  of  the  furnace,  flues  and  combustion 
chamber,  up  to  the  plane  of  first  entry  into  the  tubes.  (A.S.M.E.  Power 
Test  Codes,  Series  1929.) 

Gage  Pressure:  Pressure  measured  from  atmospheric  pressure  as  a 
base.  Gage  pressure  may  be  indicated  by  a  manometer  which  has  one  leg 
connected  to  the  pressure  source  and  the  other  exposed  to  atmospheric 
pressure. 

Grate  Area :  The  area  of  the  grate  surface,  measured  in  square  feet, 
to  be  used  in  estimating  the  rate  of  burning  fuel.  This  area  is  construed 
to  mean  the  area  measured  in  the  plane  of  the  top  surface  of  the  grate, 
except  that  with  special  furnaces,  such  as  those  having  magazine  feed,  or 
special  shapes,  the  grate  area  shall  be  the  mean  area  of  the  active  part  of 
the  fuel  bed  taken  perpendicular  to  the  path  of  the  gases  through  it. 
For  furnaces  having  a  secondary  grate,  such  as  those  in  double-grate 
down-draft  boilers,  the  effective  area  shall  be  taken  as  the  area  of  the 
upper  grate  plus  one-eighth  of  the  area  of  the  lower  grate,  both  areas 
being  estimated  as  defined  above.  (A.S.H.V.E.  Standard  and  Short 
Form  Heat  Balance  Codes  for  Testing  Low-Pressure  Steam  Heating 
Solid  Fuel  Boilers.) 

Gravity  Warm  Air  Heating  System:    See  Warm  Air  Heating  System. 

Grille:  A  perforated  covering  for  an  air  inlet  or  outlet  usually  made 
of  wire  screen,  pressed  steel,  cast-iron  or  plaster.  Grilles  may  be  plain 
or  ornamental. 

Heat:  A  form  of  energy  generated  by  the  transformation^  some  other 
form  of  energy,  as  by  combustion,  chemical  action,  or  friction.  Accord- 
ing to  the  molecular  theory,  heat  consists  of  the  kinetic  and  potential 
energy  of  the  .molecules  of  a  substance.  The  addition  of  heat  energy  to  a 
body  increases  the  temperature  or  the  kinetic  energy  of  motion  of  its 
molecules  (sensible  heat}  or  increases  their  potential  energy  of  position  but 
does  not  increase  the  temperature,  as  when  melting  or  boiling  occurs 
(latent  heat). 

Heat  Capacity :  The  amount  of  heat  (Btu  or. calories)  required  to 
raise  the  temperature  of  a  body  of  any  mass  and  variety  of  parts  one 
degree  (Fahrenheit  or  centigrade).  This  will  depend  on  the  masses  and 
specific  heats  of  the  various  parts  of  the  body. 

Therefore 

$  »  m\  Si  -f-  *»z  $1  4-  Wa  5B .  .  *  .  etc. 

where 

S  is  the  heat  capacity  and  m^  m*  #*»,  and  s\,  5j,  s\  stand  for  the  masses  and  cor- 
responding specific  heats  of  the  parts,  respectively. 

Heating  Medium:  A  substance  such  as  water,  steam,  air,  electricity 

690 


CHAPTER  41 — TERMINOLOGY 


or  furnace  gas  used  to  convey  heat  from  the  boiler,  furnace  or  other  source 
of  heat  or  energy  to  the  heating  unit  from  which  the  heat  is  dissipated. 

Heating  Surface:  The  exterior  surface  of  a  heating  unit.  Extended 
heating  surface  (or  extended  surface):  Heating  surface  having  air  on  both 
sides  and  heated  by  conduction  from  the  prime  surface.  Prime  Surface: 
Heating  surface  having  the  heating  medium  on  one  side  and  air  (or 
extended  surface)  on  the  other.  (See  also  Boiler  Heating  Surface.) 

Heat  of  the  Liquid :  The  sensible  heat  of  a  mass  of  liquid  above  an 
arbitrary  zero. 

Horsepower :  A  unit  to  indicate  the  time  rate  of  doing  work  equal  to 
550  ft-lb  per  second  or  33,000  ft-lb  per  minute.  (One  horsepower  = 
745.8  watts.  In  practice  this  is  considered  746  watts.) 

Hot  Water  Heating  System:  A  heating  system  in  which  water  is 
used  as  the  medium  by  which  heat  is  carried  through  pipes  from  the  boiler 
to  the  heating  units. 

Humidify:  To  add  water  vapor  to  the  atmosphere;  to  add  water 
vapor  or  moisture  to  any  material. 

Humidity:  The  water  vapor  mixed  with  dry  air  in  the  atmosphere. 
Absolute  humidity  refers  to  the  weight  of  water  vapor  per  unit  volume  of 
space  occupied,  expressed  in  grains  or  pounds  per  cubic  foot.  Specific 
humidity  refers  to  the  weight  of  water  vapor  in  pounds  carried  by  one  Ib  of 
dry  air.  Relative  humidity  is  a  ratio,  usually  expressed  in  per  cent,  used  to 
indicate  the  degree  of  saturation  existing  in  any  given  space  resulting 
from  the  water  vapor  present  in  that  space.  Relative  humidity  is  either 
the  ratio  of  the  actual  partial  pressure  of  the  water  vapor  in  the  air  to  the 
saturation  pressure  at  the  dry-bulb  temperature,  or  the  ratio  of  the  actual 
density  of  the  vapor  to  the  density  of  saturated  vapor  at  the  dry-bulb 
temperature.  The  presence  of  air  or  other  gases  in  the  same  space  at  the 
same  time  has  nothing  to  do  with  the  relative  humidity  of  the  space. 

Htimidis tat :  A  regulatory  device,  actuated  by  changes  in  humidity, 
used  for  the  control  of  humidity. 

Hygrostat:  Same  as  humidistat. 

Inch  of  Water:  A  measure  of  pressure  which  refers  to  the  difference 
in  the  heights  of  the  legs  of  a  water  filled  manometer. 

Insulation  (heat):  A  material  having  a  relatively  high  heat-resistance 
per  unit  of  thickness. 

Isobaric:  An  adjective  used  to  indicate  a  change  taking  place  at  con- 
stant pressure. 

Isothermal:  An  adjective  used  to  indicate  a  change  taking  place  at 
constant  temperature. 

Latent  Heat :  See  Heat. 

Laws  of  Thermodynamics :  The  first  law  states  that  the  total  energy 
of  an  isolated  system  remains  constant  and  cannot  be  increased  or  dimini- 
shed by  any  physical  process  whatever.  The  second  law  states  that  no 
change  in  a  system  of  bodies  that  takes  place  of  itself  can  increase  the 
available  energy  of  a  system. 

Manometer:  An  instrument  for  measuring  pressures;  essentially  a 
U-tube  partially  filled  with  a  liquid,  usually  water,  mercury,  or  a  light 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

oil,  so  the  amount  of  displacement  of  the  liquid  indicates  the  pressure 
Toeing  exerted  on  the  instrument. 

Mass:  The  quantity  of  matter,  in  pounds,  to  which  the  unit  of  force 
(one  pound)  will  give  an  acceleration  of  one  foot  per  second  per  second. 

W 

m  =  — -. 
g 

Mb,  Mbh1:  Symbols  which  represent,  respectively,  1000  Btu  and 
1000  Btu  per  hour. 

Mechanical  Equivalent  of  Heat:  The  mechanical  energy  necessary 
to  produce  1  Btu  of  heat  energy.  J  =  777.5  ft-lb. 

Micron:  A  unit  of  length,  the  thousandth  part  of  one  millimeter  or 
the  millionth  of  a  meter. 

Mol:  The  unit  of  weight  for  gases.  It  is  defined  as  m  Ib  where  m 
denotes  the  molecular  weight  of  a  gas.  For  any  gas  the  volume  of 
1  mol  at  32  F  and  standard  atmospheric  pressure  is  358.65  cu  ft  and  the 
weight  of  a  cubic  foot  is  0.002788  m  Ib. 

Neutral  Zone:  The  level  within  a  room  or  building  at  which  the 
pressure  is  exactly  equal  to  the  outside  barometric  pressure. 

One-Pipe  Supply  Riser  (steam):  A  pipe  which  carries  steam  upward 
to  a  heating  unit  and  which  also  carries  the  condensation  from  the  heating 
unit  in  a  direction  opposite  to  the  steam  flow. 

One-Pipe  System  (hot  water] :  A  hot  water  system  in  which  the  water 
flows  through  more  than  one  heating  unit  before  it  returns  to  the  boiler ; 
consequently,  the  heating  units  farthest  from  the  boiler  are  supplied 
with  cooler  water  than  those  near  the  boiler  in  the  same  circuit. 

One-Pipe  System  (steam) :  A  steam  heating  system  consisting  of  a 
main  circuit  in  which  the  steam  and  condensate  flow  in  the  same  pipe, 
usually  in  opposite  directions.  Ordinarily  to  each  heating  unit  there  is 
but  one  connection  which  must  serve  as  both  the  supply  and  the  return, 
although  separate  supply  and  return  connections  may  be  used. 

Overhead  System:  Any  steam  or  hot  water  system  in  which  the 
supply  main  is  above  the  heating  units.  With  a  steam  system  the  return 
must  be  below  the  heating  units;  with  a  water  system,  the  return  may 
be  above  the  heating  units. 

Panel  Radiator:  A  heating  unit  placed  on  or  flush  with  a  flat  wall 
•surface  and  intended  to  function  essentially  as  a  radiator* 

Panel  Warming :  A  method  of  heating  involving  the  installation  of 
the  heating  units  (pipe  coils)  within  the  wall,  floor  or  ceiling  of  the  room, 
so  that  the  heating  process  takes  place  mainly  by  radiation  from  the  wall, 
floor  or  ceiling  surfaces  to  the  objects  in  the  room. 

Plenum  Chamber:  An  air  compartment  maintained  under  pressure 
and  connected  to  one  or  more  distributing  ducts, 

Potentiometer:  An  instrument  for  measuring  or  comparing  small 
electromotive  forces. 

Power:  The  rate  of  performing  work,  expressed  in  units  of  horse- 
power, one  of  which  is  equal  to  550  ft-lb  of  work  per  second,  or  33,000  ft-lb 
per  minute. 

Tfas*  symbols  wert  approved  by  the  A,S.H,V.E,»  June,  1933. 

692 


CHAPTER  41 — TERMINOLOGY 


Prime  Surface :  See  Heating  Surface. 

Psychrometer:  An  instrument  for  ascertaining  the  humidity  or 
hygrometric  state  of  the  atmosphere.  Psychrometric:  Pertaining  to 
psychrometry  or  the  state  of  the  atmosphere  as  to  moisture.  Psychro- 
metry:  The  branch  of  physics  that  treats  of  the  measurement  of  degree  of 
moisture,  especially  the  moisture  mixed  with  the  air. 

Pyrometer:    An  instrument  for  measuring  high  temperatures. 

Radiation:    The  transmission  of  heat  through  space  by  wave  motion. 

Radiator:  A  heating  unit  exposed  to  view  within  the  room  or  space  to 
be  heated.  A  radiator  transfers  heat  by  radiation  to  objects  "it  can  see" 
and  by  conduction  to  the  surrounding  air  which  in  turn  is  circulated  by 
natural  convection ;  a  so-called  radiator  is  also  a  convector  but  the  single 
term  radiator  has  been  established  by  long  usage.  Concealed  Radiator: 
See  Convector. 

Recessed  Radiator:  A  heating  unit  set  back  into  a  wall  recess  but 
not  enclosed. 

Refrigerant:  A  substance  which  produces  a  refrigerating  effect  by  its- 
absorption  of  heat  while  expanding  or  vaporizing. 

Register :    A  grille  with  a  built-in  muitiblade  damper  or  shutter. 

Relative  Humidity:  See  Humidity:  see  also  discussion  of  relative- 
humidity  in  Chapter  1. 

Return  Mains :  The  pipes  which  return  the  heating  medium  from  the 
heating  units  to  the  source  of  heat  supply. 

Reversed-Return  System  (hot  water):  A  hot  water  heating  system* 
in  which  the  water  from  several  heating  units  is  returned  along  paths 
arranged  so  that  all  circuits  composing  the  system  or  composing  a  major- 
subdivision  of  the  system  are  practically  of  equal  length. 

Roof  Ventilator:  A  device  placed  on  the  roof  of  a  building  to  permit 
egress  of  air. 

Saturated  Air:  Air  containing  as  much  water  vapor  as  it  can  hold 
without  any  condensing  out;  in  saturated  air,  the  partial  pressure  of  the 
water  vapor  is  equal  to  the  vapor  pressure  of  water  at  the-  existing  tem- 
perature. 

Sensible  Heat:  See  Heat. 

Smoke:  Carbon  or  soot  particles  less  than  0.1  micron  in^size  which 
result  from  the  incomplete  combustion  of  carbonaceous  materials  such  as 
coal,  oil,  tar,  and  tobacco. 

Smokeless  Arch:  An  inverted  baffle  placed  in  an  up-draft  furnace 
toward  the  rear  to  aid  in  mixing  the  gases  of  combustion  and  thereby  to 
reduce  the  smoke  produced. 

Specific  Gravity:  The  ratio  of  the  weight  of  a  body  to  the  weight  of 
an  equal  volume  of  water  at  some  standard  temperature,  usually  39.2  F. 

Specific  Heat:  The  quantity  of  heat,  expressed  in  Btu,  required  to 
raise  the  temperature  of  1  Ib  of  a  substance  1  F. 

Specific  Volume:     The  volume,  expressed  in  cu  ft,  of  one  pound  of 

t  1         v 

a  substance,   v  =  —  •  «=  ^7. 
p        W 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

Split  System:  A  system  in  which  the  heating  and  ventilating  are 
accomplished  by  means  of  radiators  or  convectors  supplemented  by 
mechanical  circulation  of  air  (heated  or  unheated)  from  a  central  point. 

Square  Foot  of  Heating  Surface  (equivalent):  Equivalent  direct 
radiation  (EDR).  By  definition,  that  amount  of  heating  surface  which 
will  give  off  240  Btu  per  hour.  The  equivalent  square  feet  of  heating 
surface  may  have  no  direct  relation  to  the  actual  surface  area, 

Stack  Height:  The  height  of  a  gravity  con  vector  between  the  bottom 
of  the  heating  unit  and  the  top  of  the  outlet  opening. 

Standard  Air:  As  defined  by  A.S.H.V.E.  codes,  standard  air  is  air 
weighing  0.07488  Ib  per  cubic  foot,  which  is  air  at  68  F  dry-bulb  and 
50  per  cent  relative  humidity  with  a  barometric  pressure  of  29.92  in.  of 
mercury.  (Most  engineering  tables  and  formulae  involving  the  weight 
of  air  are  based  on  air  weighing  0.07495  Ib  per  cubic  foot,  which  is  dry  air 
at  70  F  dry-bulb  with  a  barometric  pressure  of  29.92  in.  of  mercury.  The 
error  involved  in  disregarding  the  difference  between  the  above  two 
weights  is  very  slight  and  in  most  instances  may  be  neglected.) 

Static  Pressure:  The  compressive  pressure  existing  in  a  fluid.  It  is 
a  measure  of  the  potential  energy  of  the  fluid. 

Steam:  Steam  is  water  vapor  which  exists  in  the  vaporous  condition 
because  sufficient  heat  has  been  added  to  the  water  to  supply  the  latent 
heat  of  evaporation  and  change  the  liquid  into  vapor.  Steam  in  contact 
with  the  water  from  which  it  has  been  generated  may  be  dry  saturated 
steam  or  wet  saturated  steam.  The  latter  contains  more  or  less  actual 
water  in  the  form  of  mist.  If  steam  is  heated,  and  the  pressure  main- 
tained the  same  as  when  it  was  vaporized,  its  temperature  will  increase 
and  it  will  become  superheated. 

Steam  Heating  System :  A  heating  system  in  which  heat  is  trans- 
ferred from  the  boiler  or  other  source  of  steam  to  the  heating  units  by 
means  of  steam  at,  above,  or  below  atmospheric  pressure. 

Steam  Trap :  A  device  for  allowing  the  passage  of  condensate  and 
preventing  the  passage  of  steam,  or  for  allowing  the  passage  of  air  as 
well  as  condensate. 

Superheated  Steam:  See  Steam. 

Supply  Mains  (steam):  The  pipes  through  which  the  steam  flows 
from  the  boiler  or  source  of  supply  to  the  run-outs  and  risers  leading  to  the 
heating  units. 

Surface  Conductance:  The  amount  of  heat  (Btu)  transmitted  by 
radiation,  conduction,  and  convection  from  a  surface  to  the  air  or  liquid 
surrounding  it,  or  vice  versa,  in  one  hour  per  square  foot  of  the  surface  for 
a  difference  in  temperature  of  1  deg  between  the  surface  and  the  sur- 
rounding air  or  liquid, 

Synthetic  Air  Chart;  A  chart  for  evaluating  the  air  conditions 
maintained  in  a  room. 

Thermal  Resistance:    The  reciprocal  of  conductance. 
Thermal  Resistivity:    The  reciprocal  of  conductivity.  * 

Thermodynamics:  The  science  which  treats  of  the  mechanical 
actions  or  relations  of  heat. 

694 


CHAPTER  41 — TERMINOLOGY 


Thermostat:  An  instrument  which  responds  to  changes  in  tempera- 
ture and  which  directly  or  indirectly  controls  the  source  of  heat  supply. 

Ton  of  Refrigeration :    The  extraction  of  12,000  Btu  per  hour. 

Ton  Day  of  Refrigeration :  The  heat  removed  by  a  ton  of  refriger- 
ation operating  for  one  day;  288,000  Btu. 

Total  Heat:  A  thermodynamic  quantity,  variously  called  heat  con- 
tent, thermal  potential,  enthalpy.  It  is  the  heat  required  per  unit  mass 
(Btu  per  Ib)  to  raise  a  given  substance  to  a  given  point  from  an  arbitrary 
datum  point.  It  is  the  sum  of  the  heat  of  the  liquid,  the  latent  heat, 
and  any  miscellaneous  heat  which  may  be  present. 

Total  Pressure:  The  sum  of  the  static  and  velocity  pressures  in  a 
fluid.  It  is  a  measure  of  the  total  energy  of  the  fluid. 

Tube  (or  Tubular)  Radiator:  A  cast-iron  heating  unit  used  as  a 
radiator  and  having  small  vertical  tubes. 

Two-Pipe  System  (steam  or  water):  A  heating  system  in  which  one 
pipe  is  used  for  the  supply  of  the  heating  medium  to  the  heating  unit  and 
another  for  the  return  of  the  heating  medium  to  the  source  of  heat 
supply.  The  essential  feature  of  a  two-pipe  system  is  that  each  heating 
unit  receives  a  direct  supply  of  the  heating  medium  which  medium  cannot 
have  served  a  preceding  heating  unit. 

Underfeed  Distribution  System  (hot  water):  A  hot  water  heating 
system  in  which  the  main  flow  pipe  is  below  the  heating  units. 

Underfeed  Stoker:  A  stoker  which  feeds  the  coal  underneath  the  fuel 
bed. 

Unit  Air  Conditioner:  A  piece  of  equipment  designed  to  provide 
simutaneous  control  of  at  least  four  of  the  seven  functions "  (page  201) 
involved  in  summer  and  winter  air  conditioning.  The  apparatus  is  com- 
pactly housed  in  a  cabinet  placed  within  or  immediately  adjacent  to  the 
rooms  served.  The  parts  comprising  a  unit  air  conditioner  are  assembled 
at  the  point  of  manufacture,  and  the  performance  of  the  assembly  is  the 
responsibility  of  the  manufacturer.  See  Chapter  12. 

Unit  Cooler:  A  cooling  device,  usually  comprising  an  extended- 
surface  element  and  a  motor-driven  fan  mounted  integrally  in  a  housing, 
located  within  or  adjacent  to  the  room  served.  Generally  no  ducts  are 
attached  to  inlet  or  outlet.  The  refrigerant  is  brought  to  the  unit  from 
an  outside  source,  and  the  fan  drives  air  over  the  cooling  element. 

Unit  Heater:  A  heating  device,  usually  comprising  an  extended- 
surface  element  or  a  gas  burner,  mounted  with  a  motor-driven  fan  in  a 
housing,  located  within  or  adjacent  to  the  room  served.  Generally,  no 
ducts  are  attached  to  inlet  or  outlet,  The  fluid  for  heating  is  brought  to 
the  unit  from  an  outside  source,  and  the  fan  drives  air  over  the  heating 
element.  Unit  heaters  are  used  primarily  in  industrial  applications. 

Unit  Ventilating-Heater :  A  ventilating  and  heating  device  com- 
prising a  motor-driven  fan,  an  extended-surface  heating  element  and 
usually  a  filter,  mounted  in  a  housing,  located  within  or  adjacent  to  the 
room  served.  Outdoor  air  is  obtained  through  a  dampered  direct  con- 
nection or  a  short  duct  from  a  nearby  wall  or  window  opening.  Provision 

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AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

for  partial  recirculation  is  usually  made.  If  a  humidifier  is  included,  such 
a  filter-equipped  device  becomes  a  winter-type  unit  conditioner.  Unit 
ventilators  are  used  primarily  for  offices,  schools,  and  places  of  public 
assembly. 

Up-Feed  System  (steam):  A  steam  heating  system  in  which  the 
supply  mains  are  below  the  level  of  the  heating  units  which  they  serve. 

Vacuum  Heating  System:  A  two-pipe  steam  heating  system  equip- 
ped with  the  necessary  accessory  apparatus  which  will  permit  operating 
the  system  below  atmospheric  pressure  when  desired. 

Vapor:  Any  substance  in  the  gaseous  state. 

Vapor  Heating  System:  A  steam  heating  system  which  operates 
under  pressures  at  or  near  atmospheric  and  which  returns  the  condensa- 
tion to  the  boiler  or  receiver  by  gravity.  Vapor  systems  have  thermo- 
static  traps  or  other  means  of  resistance  on  the  return  ends  of  the  heating 
units  for  preventing  steam  from  entering  the  return  mains;  they  also  have 
a  pressure-equalizing  and  air-eliminating  device  at  the  end  of  the  dry 
return.  Direct  Vent  Vapor  System:  A  vapor  heating  system  with  air 
valves  which  do  not  permit  re-entry  of  air. 

Vapor  Pressure:  The  equilibrium  pressure  exerted  by  a  vapor  in 
contact  with  its  liquid. 

Velocity:  The  time  rate  of  motion  of  a  body  in  a  fixed  direction.  In 
the  fps  system  it  is  expressed  in  units  of  one  foot  per  second.  V  =  ™. 

Velocity  Pressure:  The  pressure  corresponding  to  the  velocity  of 
flow.  It  is  a  measure  of  the  kinetic  energy  of  the  fluid. 

Ventilation:  The  process  of  supplying  or  removing  air  by  natural  or 
mechanical  means,  to  or  from  any  space.  Such  air  may  or  may  not  have 
been  conditioned.  (See  Air  Conditioning.) 

Warm  Air  Heating  System :  A  warm  air  heating  plant  consists  of  a 
heating  unit  (fuel-burning  furnace)  enclosed  in  a  casing,  from  which  the 
heated  air  is  distributed  to  the  various  rooms  of  the  building  through 
ducts.  If  the  motive  head  producing  flow  depends  on  the  difference  in 
weight  between  the  heated  air  leaving  the  casing  and  the  cooler  air 
entering  the  bottom  of  the  casing,  it  is  termed  a  gravity  system.  A  booster 
fan  may,  however,  be  used  in  conjunction  with  a  gravity-designed 
system.  If  a  fan  is  used  to  produce  circulation  and  the  system  is  designed 
especially  for  fan  circulation,  it  is  termed  a  fan  furnace  system  or  a 
central  fan  furnace  system.  A  fan  furnace  system  may  include  air  washers 
and  filters. 

Wet-Bulb  Temperature:  The  lowest  temperature  which  a  water 
wetted  body  will  attain  when  exposed  to  an  air  current.  This  is  the 
temperature  of  adiabatic  saturation, 

Wet  Return:  That  part  of  a  return  main  of  a  steam  heating  system 
which  is  filled  with  water  of  condensation.  The  wet  return  usually  is 
below  the  level  of  the  water  line  in  the  boiler,  although  not  necessarily  so. 


CHAPTER  41 — TERMINOLOGY 


ABBREVIATIONS2 

Absolute. abs 

Acceleration,  due  to  gravity g 

Acceleration,  linear.— a 

Air  horsepower air  hp 

Alternating-current  (as  adjective) a-c 

Ampere amp 

Ampere-hour. amp-hr 

Area... A 

Atmosphere.— atm 

Average..... avg 

Avoirdupois avdp 

Barometer. bar. 

Boiler  pressure bp 

Boiling  point bp 

Brake  horsepower bhp 

Brake  horsepower-hour bhp-hr 

British  thermal  unit Btu 

Calorie -• cal 

Centigram eg 

Centimeter cm 

Centimeter-gram-second  (system) . cgs 

Change  in  specific  volume  during  vaporization z>fg 

Cubic cu 

Cubic  foot.... - cu  ft 

Cubic  feet  per  minute -- cfm 

Cubic  feet  per  second cfs 

Decibel db 

Degree3 .„ - <teg  or  ° 

Degree  centigrade - C 

Degree  Fahrenheit F 

Degree  Kelvin - K 

Degree  Reaumur - R 

Density,  Weight  per  unit  volume,  Specific  weight A  or  p  (rho) 

1 

P  =  *V 

Diameter D  or  diam 

Direct-current  (as  adjective) d-c 

Distance,  linear $ 

Dry  saturated  vapor,  Dry  saturated  gas  at  saturation  pressure  and  temperature, 

Vapor  in  contact  with  liquid '.....Subscript  g 

Entropy  (The  capital  should  be  used  for  any  weight,  and  the  small  letter  for  unit 

weight.) -S  or  ^ 

Feet  per  minute fPm 

Feet  per  second ....fps 

Foot , - y:«it 

Foot-pound *t;lt> 

Foot-pound-second  (system) - fps 

Force,  total  load f 

Freezing  point - • *P 

Gallon - Sal 

Gallons  per  minute £Pm 

Gallons  per  second - SPS 

Gram 3 

Gram-calorie ^ - g"cal 

*From  compilations  of  abbreviations  approved  by  the  American  Standards  Association,  Z,  10  a,  c,  ft  and 
i,  As  a  general  rule  the  period  is  omitted  in  all  abbreviations  except  where  the  omission  results  in  the 
formation  of  an  English  word. 

•It  is  recommended  that  the  abbreviation  for  the  temperature  scale,  F,  C,  K,  be  included  in  expressions 
for  numerical  temperatures  but,  wherever  feasible,  the  abbreviation  for  degree  be  omitted;  as  68  F. 

697 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

Head  ....................................................................................................................................  H  or  h 

Heat  content,  Total  heat,  Enthalpy.    (The  capital  should  be  used  for  any  weight 

and  the  small  letter  for  unit  weight.)  ......................................................................  H  or  h 

Heat  content  of  saturated  liquid,  Total  heat  of  saturated  liquid,  Enthalpy  of 

saturated  liquid,  sometimes  called  heat  of  the  liquid  ................................................  hf 

Heat  content  of  dry  saturated  vapor,  Total  heat  of  dry  saturated  vapor,  Enthalpy 

of  dry  saturated  vapor  ......................................................................................................  h& 

Heat  of  vaporization  at  constant  pressure  ..................................................................  L  or  hfs 

Horsepower  ................................................................................................................................  hp 

Horsepower-hour™  ...............................................................................................................  hp-hr 

Inch  ........................................................................................................................................  v..in. 

Inch-pound  ............................................................................................................................  in-lb 

Indicated  horsepower.  ...............  j  ........................................................................................  .....ihp 

Indicated  horsepower-hour..—  ...........................................................................................  ihp-hr 

Internal  energy,  Intrinsic  energy.    (The  capital  should  be  used  for  any  weight  and 

the  small  letter  for  unit  weight.)  ..............................................................................  U  or  u 

Kilogram  ....................................................................................................................................  kg 

Kilowatt  ....................................................................................................................................  kw 

Kilowatthour  ........................................................................................................................  kwhr 

Length  of  path  of  heat  flow,  thickness  ....................................................................................  L 

Load,  total  ___  .............................................................................................................................  W 

Mass  ............................................................................................................................................  tn 

Mechanical  efficiency  ................................................................................................................  em 

Mechanical  equivalent  of  heat  ..................................................................................................  / 

Melting  point  ................................................................................................  '.  ...........................  mp 

Meter  ..........................................................................................................................................  m 

Micron  ................  r  ...............................................................................................................  &  (niu) 

Miles  per  hour  ........................................................................................................................  mph 

Minute  ......................................................................................................................................  mm 

Molecular  weight  ............................................................................................................  mol.  wt 

Mol  ................................................................................................................................  .  ...........  mol 

Ounce  .........................................................................................................................................  .02 

Power,  Horsepower,  Work  per  unit  time  ___  .  ..........................................................................  P 

Pressure,  Absolute  pressure,  Gage  pressure,  Force  per  unit  area  ................  .  .......................  p 

Quantity  (total)  of  fluid,  water,  gas,  heat;  Quantity  by  volume;  Total  quantity 

of  heat  transferred  .............................................................  .  ...............................  ,  ..............  Q 

Quality  of  steam,  Pounds  of  dry  steam  per  pound  of  mixture  .........................  -  ...........  .  .....  ...» 

Revolutions  per  minute  ............................................................  .  ...........................................  rprn 

Saturated  liquid  at  saturation  pressure  and  temperature,  Liquid  in  contact 

with  vapor  ..........................................................................................................  Subscript  f 

Specific  gravity  ..........................................................................................................  ,  .........  sp  gr 

Specific  neat  ...........................................................................................................  ,...»,.sp  ht  or  c 

Specific  heat  at  constant  pressure  ...............................................................  .  ..........................  £p 

Specific  heat  at  constant  volume  .................................................................  .  ..........................  cv  ' 

Specific  volume,  Volume  per  unit  weight,  Volume  per  unit  mass  ..........  ..  ............................  v 

Square  foot...  ...................................  .  ...............................  ,  ......  ,  ...............  .„,.  .......  ,........»„„  ......  sq  ft 

Square  inch  ..........................................................  „  .............  .  .....  „„  .....  „  ......................  ...  ......  ....sq  in. 

Temperature  (ordinary)  F  or  C.   (Theta  is  used  preferably  only  when  t  is  used  for 

Time  in  the  same  discussion.)...,.  ..................  „„.„  ...................  .  ....................  /  or  6  (theta) 

Temperature  (absolute)  F  abs  or  K,    (Capital  theta  is  used  preferably  only  when 

small  theta  is  used  for  ordinary  temperature,)  ...........  ,  .........  „„  ____  T  or  €>  (capital  tketa) 

Thermal  conductance4  (heat  transferred  per  unit  time  per  degree),,  .........................  ....  .....  C 

r        L       j¥       _£_ 

°  "  ~R    "    L    "  ti  ~  it 

Thermal  conductance  per  unit  area,  Unit  conductance  (heat  transferred  per 

unit  time  per  unit  area  per  degree)  ......................................................  .  .........  .......  .....  Ca 


Cfl   «  JL   m  JL  _ 

°a        A        RA       Afa-h)        L 

<Terms  endinff  ivity  designate  properties  independent  of  size  or  shape,  sometimes  called  spsctyt  proper- 
te$.  Examples  are  —  conductivity  and  resistivity.  Terms  ending  attcs  designate  quantities  depending 
lot  only  on  the  material,  but  also  upon  size  and  shape,  sometimes  called  total  qv&ntitiw*  Example*  are  — 
:onductance  and  transmittance.  Terms  ending  ion  designate  rate  of  heat  transfer.  Examples  are*—  con- 
luction  and  transmission. 

698 


CHAPTER  41 — TERMINOLOGY 


Thermal  conductivity   (heat  transferred   per  unit  time  per  unit  area,  and  per 
degree  per  unit  length) 


JL. 
A 


Of  \ 
i  —  tz) 

L 

Surface  coefficient  of  heat  transfer,  Film  coefficient  of  heat  transfer,  Individual 
coefficient  of  heat  transfer  (heat  transferred  per  unit  time  per  unit  area 
per  degree) ./ 


f 
7 


-  /* 

(In  general  /  is  not  equal  to  k/L,  where  L  is  the  actual  thickness  of  the  fluid  film.) 

Over-all  coefficient  of  heat  transfer,  Thermal  transmittance  per  unit  area  (heat 

transferred  per  unit  time  per  unit  area  per  degree  over-all)  ......................................  U 


Thermal  transmission  (heat  transferred  per  unit  time)  ........................................................  q 

•  -9- 

Thermal  resistance  (degrees  per  unit  of  heat  transferred  per  unit  time)  ..........  ,  ..............  R 


Thermal  resistivity  .........................................................................................  .  ........................  l/k 

Vaporization  values  at  constant  pressure,  Differences  between  values  for  saturated 

vapor  and  saturated  liquid  at  the  same  pressure  ..................  ,  .......................  Subscript  fg 

Velocity.  .......................................................................................................................................  v 

Volume  (total)  ................................................................................................................  .•  ...........  V 

Volume  per  unit  time,  Rate  at  which  quantity  of  material  passes  through  a 

machine,  Quantity  of  heat  per  unit  time,  Quantity  of  heat  per  unit  weight  ............  g 

Watt  .........................  »  ..................................................................................................................  w 

Watthour....,  ............................................  ...  .........  .  ....................  .  ................................  ,  ..............  whr 

Weight  of  a  major  item,  Total  weight  ......................  >  ............................  .  ..................  .  ...........  W 

Weight  rate,  Weight  per  unit  of  power,  Weight  per  unit  of  time  ......................................  w 

Work  (total)  .....................................................................................................  .  .......  .  ...............  W 

CONVERSION  EQUATIONS 

Fahrenheit  degrees  «  9/5  centigrade  degrees  +  32, 

Centigrade  degrees  =»  5/9  (Fahrenheit  degrees  —  32). 

Absolute  temperature,  expressed  in  Fahrenheit  degrees  *»  Fahrenheit  degrees  + 
459,6,  In  heating  and  ventilating  work,  460  is  usually  used. 

Absolute  temperature,  expressed  in  centigrade  degrees  »  centigrade  degrees  -f- 
273.1. 

Power,  Heat,  and  Work 

I  ton  Deration  - 

Latent  heat  of  ice  »  143.33  Btu  per  pound 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


1  Btu 

1  watthour 

1  mean  calorie 

1  kilowatt  (1000  watts) 

1  horsepower 

1  boiler  horsepower 

Weight  and  Volume 

1  gal  (U.  S.) 

1  British  or  Imperial  gallon 

1  cu  ft 

1  cu  ft  water  at  60  F 
1  cu  ft  water  at  212  F 
1  gal  water  at  60  F 
1  gal  water  at  212  F 

1  Jb  (avdp) 

1  bushel 
1  short  ton 
1  long  ton 

Pressure 

1  lb  per  square  inch 

1  oz  per  square  inch 
1  atmosphere 

1  in.  water  at  62  F 
1  ft  water  at  62  F 

1  in.  mercury  at  62  F 


f  777.5  ft-lb 

\  0.293  watthours 

[  252.02  mean  calories 

f  2,655.2  ft-lb 

I  3.415  Btu 

|  3600  joules 

{  860.648  mean  calories 

0.003968  Btu 

3.085  ft-lb 

0.0011619  watthours 
f  1.3405  horsepower 
\  56.92  Btu  per  minute 
[  44,252.7  ft-lb  per  minute 
[  0.746  kilowatt 
J  42.44  Btu  per  minute 
1  33,000  ft-lb  per  minute 
[  550  ft-lb  per  second 
33,471.9  Btu  per  hour 


f  231  cu  in. 

\  0.13368  cu  ft 

277.274  cu  in. 
7.4805  gal 
1728  cu  in. 

62.37  lb 

59.76  lb 

8.34  lb 

7.99  lb 

/  16  oz 

\  7000  grains 

1.244  cu  ft 

2000  lb 

2240  lb 


144  lb  per  square  foot 
2.0416  in.  mercury  at  62  F 
2.309  ft  water  at  62  F 
27.71  in.  water  at  62  F 
0.1276  in.  mercury  at  62  F 
1.732  in.  water  at  62  F 
14.7  lb  per  square  inch 
2116.3  lb  per  square  foot 
33.974  ft  water  at  62  F 
30  in.  mercury  at  62  F 
29.921  in.  mercury  at  32  F 
0.03609  lb  per  square  inch 
0.5774  02  per  square  inch 
5.196  lb  per  square  foot 
I  0.433  lb  per  square  inch 
\  62,355  lb  per  square  foot 

10.491  lb  per  square  inch 
7.86  oz  per  square  inch 
U31  ft  water  at  62  F 
13.57  in.  water  at  62  F 


700 


CHAPTER  41 — TERMINOLOGY 


Metric  Units 

1  cm 

lin. 

1  m 

1ft 

1  sq  cm 

1  sq  in. 

1  sq  m 

1  sqft 

1  cu  cm 

1  cu  in. 

1  cu  m 

leu  ft 

1  liter 

1  kg 

lib 

1  metric  ton 

1  gram 

1  kilometer  per  hour 

'  1  gram  per  square  centimeter 
1  kg  per  square  centimeter  (metric  atmosphere) 
1  gram  per  cubic  centimeter 
1  dyne 
I  joule 

1  metric  horsepower 

1  kilogram-calorie  (large  calorie) 

1  kilogram-calorie  per  kilogram 

1  gram-calorie  per  square  centimeter 

1  gram-calorie  per  square  centimeter  per  centi- 
meter 

1  gram-calorie  per  second  per  square  centimeter 
for  a  temperature  graduation  of  1  deg  C  per 
centimeter 


0.3937  in, 

2.54  cm 

3.281  ft 

0.3048  m 

0.155  sq  in. 

6.45  sq  cm 

10.765  sq  ft 

0.0929  sq  m 

0.061  cu  in. 

16.39  cu  cm 

35.32  cu  ft 

0.0283  cu  m 

1000  cu  cm  =  0.264  gal 

•  2.2046  Ib 

'  0.4536  kg 

'  2205  Ib  (avdp) 

>  980.59  dynes  =  0.002205  Ib 

'  0.6214  mph 

/  0.0290  in.  mercury,  at  0  deg  C 
1  \  0.394  in.  water,  at  15  C 
'  14.22  Ib  per  square  inch 

/  0.03614  Ib  per  cubic  inch 
1  \  62.43  Ib  per  cubic  foot 
'  0.00007233  poundals 

/  10,000,000  ergs 
!  \  0.73767  ft-lb 

/  75  kg-m  per  second 
:  \  0.986  hp  (U.  S.) 

f  1000  gram-calories  (small 
'  \      calorie) 

(  3.97  Btu 

•  1.8  Btu  per  pound 

•  3.687  Btu  per  square  foot 

'  >  1.451  Btu  per  square  foot  per  inch 

[2903  Btu  per  hour  per  square  foot 

M  for  a  temperature  graduation  of 

(  1  deg  F  per  inch  of  thickness. 


SYMBOLS  FOR  HEATING  AND  VENTILATING  DRAWINGS5 

1.  The  objects  of  this  standard  set  of  symbols  are  to  insure  the  correct  interpretation 
of  drawings  and  to  conserve  drafting  room  time  by  establishing  simple  and  unmistakable 
symbols  for  the  component  parts  of  the  heating  and  ventilating  systems.    In  preparing 
the  Hat  of  symbols  an  effort  has  been  made  to  follow  existing  practice  in  so  far  as  possible 
but  the  list  cannot  be  expected  to  match  exactly  the  existing  practice  of  every  drafting 
room. 

2.  Simplicity,  ease  of  execution  and  unmistakable  identification  were  carefully  con- 
sidered in  selecting  the  symbols.    Uncommon  fittings  and  appliances  such  as  vacuum 
pumps,  separators,  etc.,  have  purposely  been  omitted  in  order  to  produce  a  list  which 
can  be  easily  remembered.    It  is  assumed  that  when  the  scale  of  the  drawing  permits, 
the  valves  and  fittings  will  be  drawn  to  scale  and  a  conventional  representation  is  then 
unnecessary, 


•From  A.S.H.V.E,  Code  of  Minimum  Requirements  for  the  Heating  and  Ventilation  of  Buildings, 
edition  of  1929, 

701 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

3.  High  pressure  steam  supply  pipe  —  - 

4.  Low  pressure  steam  supply  pipe  ' 

5.  Hot  water  pipe  —  flow  "***  ---- 

6.  Return  pipe  —  steam  or  water  -------  "-• 


7.  Air  vent  line  

8.  Flanges  — :ff— 

9.  Screwed  union  — |J|| — 

10.  Elbow  "VH 

11.  Elbow — looking  up  Of—* 

12.  Elbow — looking  down  @* 

13.  Tee  -fr~ 

14.  Tee— looking  up  -H©h- 

15.  Tee — looking  down  Ol 

16.  Gate  valve  ill  • 

17.  Globe  valve 

18.  Angle  valve 

19.  Angle  valve — stem  perpendicular 

20.  Lock  shield  valve  ^m\    H^l 

21.  Check  valve 

22.  Reducing  valve 


23.  Diaphragm  valve 

702 


CHAPTER  41 — TERMINOLOGY 


24.  Diaphragm  valve — stem  perpendicular 

25.  Thermostat 

26.  Radiator  trap — elevation 

27.  Radiator  trap — plan 

28.  Expansion  joint 

29.  Column  radiator — plan 

30.  Column  radiator — elevation 

31.  Wall  radiator— plan 

32.  Wall  radiator — elevation 

33.  Pipe  coil — plan 

34.  Pipe  coil — elevation 

35.  Indirect  radiator — plan 

36.  Indirect  radiator — elevation 

37.  Supply  duct — section 

38.  Exhaust  duct—- section 

39.  Butterfly  damper — plan  (or  elevation) 

40.  Butterfly  damper — elevation  (or  plan) 

41.  Deflecting  damper— square  pipe 

42.  Vanes 

43.  Air  supply  outlet 

44.  Exhaust  outlet 

708 


Oh 


-9 

-I® 


f=l 

en 


n 

U iJ 


\z\ 


n 

0 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


A.S.H.V.E.  CODES 

The  following  codes  and  standards  relating  to  the  design,  installation, 
testing,  rating,  and  maintenance  of  materials  and  equipment  used  for  the 
heating  and  ventilation  of  buildings,  have  been  adopted  by  the  AMERICAN 
SOCIETY  OF  HEATING  AND  VENTILATING  ENGINEERS: 


SUBJECT 

TITLE 

WHEN  ADOPTED 

REFERENCE 

Air 
Cleaning 

A.S.H.V.E.  Standard  Code  for 
Testing  and  Rating  Air  Clean- 

June  19o4 

A.S.H.V.E.  Reprint 

Devices 

ing  Devices  Used  in  General 
Ventilation  Work 

Air  purity 

Synthetic  Air  Chart 

June,  1917 

A.S.H.V.E. 
TRANSACTIONS, 
Vol.  23,  p.  607,  and 
THE  GUIDE,  1931 

Standard  and  Short-Form  Heat 

A.S.H.V.E. 

Boilers 

(testing) 

Balance  Codes  for  Testing  Low 
Pressure  Steam  Heating  Solid 
Fuel  Boilers  (Codes  1  and  2) 

June,  1929 

TRANSACTIONS, 
Vol.  35,  1929 

A  S.H  V.E.  Performance  Test 

A  S.H.V.E. 

Boilers 

(testing) 

Code  for  Steam  Heating  Solid 
Fuel  Boilers  (Code  3)a 

June,  1929 

TRANSACTIONS, 
Vol.  35,  1929 

Boilers  — 

A.S.H  V  E.  Standard  Code  for 

A  S  H.V.E. 

Oil  Fuel 

(testing) 

Testing  Steam  Heating  Boilers 
Burning  Oil  Fuel 

June,  1932 

TRANSACTIONS, 
Vol.  37,  1931 

ID     '1 

A.S.H.V.E.  Standard  Code  for 

January,  1929 

A.S.H.V.E. 

Boilers 
(rating) 

Rating  Steam   Heating  Solid 
Fuel  Hand  Fired  Boilers 

Revised 
April,  1930 

TRANSACTIONS, 
Vol.  30,  1930,  p.  42 

[    Concealed 
^  Gravity 

A.S.H.V.E.  Standard  Code  for 
Testing  and  Rating  Concealed 

June,  1934 

A.S.H.V.E.  Reprint 

Type 
Radiation 

Gravity  Type  Radiation  (Hot 
Water  Section) 

Con  vectors 

A.S.H.V.E.  Standard  Code  for 
Testing  and  Rating  Concealed 
Gravity  Type   Radiation 
(Steam  Code) 

January,  1931 

A.S.H.V.E. 

TRANSACTIONS, 
Vol.  37,  1931,  p.  367 

Ethics 

Code  of  Ethics  for  Engineers 

January,  1922 

A.S.H.V.E. 
TRANSACTIONS, 
Vol.  28,  1922,  p.  6 
(See  frontispiece 
THE  GUIDE,  1935) 

Fans 

Standard  Test  Code  for  Disc 
and  Propeller  Fans,  Centrifugal 
Fans  and  Blowers 

May,  1923. 
Revised 
June,  1931 

A.S.H.V.E. 
TRANSACTIONS, 
Vol.  29,  1923, 
p.  407>> 

Garages 

Code  for   Heating  and   Ven- 
tilating Garages 

June,  1929 

A.S.H.V.E.  TRANS- 
ACTIONS,     Vol.    35, 
1929,  p.  355 

^Originally  adopted  by  the  National  Boiltr  and  Radiator  Manufactures  Association, 
o,  see  Hwtfag,  Piping  and  Air  Conditioning,  August,  1931,  p,  743, 

704 


CHAPTER  41 — TERMINOLOGY 


SUBJECT 

TITLE 

WHEN  ADOPTED 

REFERENCE 

Heat 
transmission 
through  walls 

Standard  Test  Code  for  Heat 
Transmission  through  Walls 

January,  1927 

A.S.H.V.E. 
TRANSACTIONS, 
Vol.  34,  1928,  p.  253- 

Minimum 
requirements 

Code   of    Minimum    Require- 
ments for  Heating  and  Ventila- 
tion of  Buildings,  Edition-1929 

June,  1925 

A.S.H.V.E. 
Codes 

Pitot  tube 

Code  for  Use  of  Pitot  Tube 

January,  1914 

A.S.H.V.E. 
TRANSACTIONS, 
Vol.  20,  1914,  p.  211 

Radiators 

Code    for    Testing    Radiators 

January,  1927 

A.S.H.V.E. 
TRANSACTIONS, 
Vol.  33,  1927,  p.  18- 

Unit  heaters 

Standard  Code  for  Testing  and 
Rating  Steam   Unit   Heatersc 

January,  1930 

A.S.H.V.E. 
TRANSACTIONS, 
Vol.  36,  1930,  p.  165- 

A  S  H.V  E  Standard  Code  for 

A  S  H  V  E. 

Unit 
Ventilators 

Testing    and    Rating    Steam 
Unit  Ventilators 

June,  1932 

TRANSACTIONS, 
Vol.  38,  1932,  p.  25 

A  S  H  V  E  Standard  Code  for 

ASH  VE  Journal- 

Vacuum 
Heating 
Pumps 

Testing    and    Rating    Return 
Line    Low    Vacuum    Heating 
Pumps 

June,  1934 

Heating,  Piping  and 
Air   Conditioning, 
March,  1934,  p.  136 

A.S.H.V.E. 

Ventilation 

Report    of    Committee    on 
Ventilation  Standards 

August,  1932 

TRANSACTIONS, 
Vol.  38,  1932,  p.  38£ 

The  following  Codes  and  Standards  have  been  endorsed  or  approved 
by  the  AMERICAN  SOCIETY  OF  HEATING  AND  VENTILATING  ENGINEERS: 


SUBJECT 

TWLB 

SPONSORED  BT 

RKFBRENCB 

Chimneys 

Standard  Ordinance  for  Chim- 
ney Construction 

National  Board  of 
Fire  Underwriters 

Chapter  14, 
THE  GUIDE,  1931 

Piping 
systems 

Identification 
of  Piping  Systems*1 

A  merican  Society 
of  Mechanical 
Engineers 

Heating^  Piping  an& 
Air  Conditioning, 
July,  1929 

Warm  air 
furnaces 

Standard  Code  Regulating  the 
Installation  of  Gravity  Warm 
Air    Furnaces    in    Residences 

National    Warm 
Air  Heating  As- 
sociation 

National  Warm  Air 
Heating  A  ssociation  , 
Columbus,  Ohio 

oAdopted  jointly  by  the  Industrial  Unit  Heater  Association,  and  the  A.S.H.V.E. 
d  Adopted  November,  1928,  Sponsored  by  (I)  American  Society  of  Mechanical  Engineers,  (2)  Nationa? 
Safety  Council. 


705 


INDEX 

THE  A.  S.  H.  V.  E.  GUIDE  1935 

Technical  Data  Section 

Chapters  1-41  and  Pages  1-705 


Titles  of  Chapters 


1.  Fundamentals    of   Heating   and    Air  21.  Industrial  Exhaust  Systems. 

Conditioning. 

22.  Fan  Systems  of  Heating. 

2.  Ventilation     and     Air     Conditioning 

Standards,  23.  Mechanical  Warm  Air  Furnace  Sys- 
tems. 

3.  Industrial  Air  Conditioning. 

24.  Gravity  Warm  Air  Furnace  Systems. 

4.  Natural  Ventilation. 

25.  Boilers. 

5.  Heat  Transmission   Coefficients   and 

Tables.  26,  Chimneys  and  Draft  Calculations, 

6   Air  Leakage  27,  Fuels  and  Combustion. 

7.  Heating  Load.  28.  Automatic  Fuel  Burning  Equipment. 

8.  Cooling  Load.  29.  Fuel  Utilization. 

9.  Central  Air  Conditioning  Systems,  30*  Radiators  and  Gravity  Convectors. 

10.  Cooling  Methods.  3L  Steam  Heating  Systems. 

11.  Humidification  and  Dehumidification.  32'  PiPinS  for  Steara  Heating  Systems. 

12.  Unit  Air  Conditioners  and  Condition-  33-  Hot¥  Water    Heating    Systems    and 

ing  Systems.  Piping, 

13.  Unit  Heaters,  Ventilators  and  Coolers.  34.  Pipe,  Fittings,  Welding. 

14.  Automatic  Control.  35.  Water  Supply  Piping. 

15.  Air  Pollution.  36.  Insulation  of  Piping. 

16.  Air  Cleaning  Devices.         ,  37.  District  Heating. 

17.  Fans  and  Motive  Power.  38.  Radiant  Heating. 

18.  Sound  Control  39.  Electrical  Heating. 

19.  Air  Distribution,  40.  Test  Methods  arid  Instruments. 

20.  Air  Duct  Design.  41.  Terminology. 


708 


INDEX 

TECHNICAL   DATA   SECTION 

(Pages  1  to  705) 

Cross  Reference  to  Subjects  in 

Chapters  1  to  41  Alphabetically 

Listed 


Air  (continued). 

Page 

A 

Page 

friction  of,  in  pipes, 

327 

Abbreviations, 

697 

impurities  in, 

84,  259 

Absolute  humidity, 

7 

size  of, 

271 

Absorption, 
as  means  of  dehumidification, 

(see  also  Regain) 
166 

ionization  of, 
leakage, 

56 
119,126 

of  solar  radiation  by  glass, 

151 

mixtures  with  water  vapor, 

10 

of  sound, 

303 

moist, 

45 

Acceleration, 

685 

motion, 

51 

Acclimatization, 
Acoustics,  acoustical, 
effect  of  humidity  on, 

38 
299 
312 

odora  in, 
optimum  conditions, 
indoors  in  summer, 
outside,  introduced, 

33 

42 
48 

treatment, 
Adiabatic  saturation, 

304 
14,  166,  184,685 

effect  on  temperature, 
fan  systems, 

54 
364,  367 

Adsorption, 
as  means  of  dehumidification, 

685 
166 

through  cracks, 
unit  air  conditioners. 

126 
212 

unit  ventilators, 

228 

Air, 

pollution, 

259 

adlabatic  saturation  of, 

14 

abatement  of, 

263 

amount  per  person, 

52 

effect  on  health, 

260 

atmospheric, 

1 

primary, 

445 

changes  of,  indoors, 
cleaning  devices, 
A,S.H,V.E.  code  for, 
ratings  of, 

33,  126 
271,  685 
272 
272 

properties  of, 
Quality, 
quantity  necessary, 
for  combustion, 

2,  4,  49 
51 

445,  454 

requirements  of, 

271 

for  ventilation, 

52,  500 

types  of. 

273 

recirculation  of, 

55 

composition  of, 

1,33 

fan  systems, 

360,  367 

density  of, 

6 

unit,  ventilators, 

229 

distribution  of, 

317 

saturated, 

1,5,9,693 

A.S.H,V,E.  standards, 
for  comfort, 

49 
51 

secondary, 
apace  conductances, 

445,  473 
94 

downward, 

321 

standard, 

694 

natural  ventilation* 

83 

still, 

41 

in  theaters, 
with  unit  ventilators. 

321 
220 

velocity, 
vitiation, 

(see  Velocity,  Air) 
83 

upward  » 

321 

volume, 

10 

dry,                                                             l,4,8,4fi 
ducta,                                      325  (see  Ducts,  Air) 

washer,                   183,  686  (see  also  Washer  ,  Air) 
cooling  towers  for,                                      187 

exfiltration, 
filtration,                            119, 
flow, 
control,  principles  of, 

119,363,479 
128,363,479,685 
49,51 
83 

humidifying  efficiency, 
operation  of, 
saturation*  efficiency, 
weight  of, 

369 
252 
184 
6 

as  cooling  method, 
diagrams, 

165,  387 
325 

Air  conditioning,         33,  49, 

72  ,085  (see  also  Air) 

formulae, 

325 

air  change  per  occupant, 

53 

into  a  hood, 

S50 

A.S.H.V.E,  standards, 

49 

natural,  measurement  of, 

86 

chemical  factors, 

33,  49 

requirements. 

84 

comfort  chart, 

44,46 

tables, 

325 

functions  of, 

201 

through  openings, 

78,83 

fundamentals  of, 

1 

709 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


Air  conditioning  (continued), 

industrial, 
apparatus  for, 
automatic  control, 
exhaust  systems, 
of  libraries, 
plants, 

process  conditioning, 
of  railroad  cars, 
temperature  differential  in, 
unit  air  conditioners, 

objective  of, 

physical  factors, 

recirculation  of  air, 

standards, 
Algae  formations, 
Alumina  system  of  adsorption, 
Aluminum  foil, 
Ammonia, 
Anemometer, 
Anthracite, 
Apartment  houses, 

hot  water  supply  to, 

stokers  suitable  for, 
Aquastat, 
Area  of: 

chimneys, 

fittings, 

grates, 

human  body  surface, 

leader  pipes, 

pipe, 

registers, 

stacks, 

wall  surfaces, 

A.S.H.V.E.  Codes  and  Standards, 
48, 88,  2 

A.S.M.E,  boiler  construction  code. 
Asbestos, 
Ash, 

cared  for  by  stokers, 

ny, 

Atmosphere,  standard, 
Atmospheric  steam  heating  system, 
Atmospheric  water  cooling  apparatus, 

design  of, 

efficiency  of, 
Atoraization, 

for  humidifying, 

of  oil, 

Audiometer, 
Automatic  control, 
Automatic  fuel  burning  equ 
Awninga, 

B 

Bab  cock's  formula  for  steam  flow, 

Baffles, 

Bananas,  ripening  of, 

Barn  ventilation, 

Barometer, 

aneroid, 

mercurial, 
Baudelot, 

chamber, 

heat  absorber, 
Bends,  expansion, 

BET,  British  equivalent  temperature. 
Blast, 
Blower,  blowers, 

standard  teat  code  for. 
Blow-through  heating  units, 
Body,  human,  surface  area, 
Boiler,  boilers, 

A,S,H.V,E,  test  codes, 

A.S.M.E.  construction  code, 

baffiles, 

capacity, 

care  during  summer, 

cleaning, 

connections, 

conversion, 

design  of, 

domestic  oil  burners, 


«0,                           Page 

Boiler,  boilers  (continued), 

Page 

65 

draft  loss  through, 

435 

72 

efficiency  of,                                         409, 

411,480 

250 

for  electric  steam  heating, 

670 

845 

fittings, 

416 

70 

gas-  fired,                                                407, 

415,  471 

195,  262 

heat  transfer  rates, 

'    400 

65,  167 

heating  surface,                                409, 

474,  686 

201 

horsepower, 

686 

lin,                             159 

installation, 

415,417 

200 

insulation, 

420 

1,48,58 

low  pressure,  construction  code, 

416 

33,49 

oil  burners, 

406,  470 

55,246,366,382,472 

operation, 

41S 

33,49 

output, 

410 

191 

performance  curves, 

413 

on,                             169 

ratings  of, 

410,475 

113 

runouts,  sizes, 

545 

188,  597 

scale  in, 

419 

678,  686 

selection  of, 

414 

443,  447  (see  also  Coal] 

settings, 

408 

troubles  with, 

418 

617 

types  of, 

405,  670 

459 

warming-up  allowance, 

412,414 

469 

water  line, 

416 

Booster, 

430,  440 

coils, 

361 

587 

fans. 

156,  402 

371,  383,  396,  414,  690 

Bourdon  tube, 

675 

52 

Boyle's  law, 

5 

391 

A91 

Brake  horsepower. 

OjiJL 

fan, 

187 

393 
393 

heat  equivalent  of, 
Branch  connections, 

138 
505 

92 

Breeching,  draft  loss  through, 

435 

ndards, 
:8,  88,  272,  284,  498,  704 
n  code,                     417 

Brine, 
British  equivalent  temperature, 
British  thermal  unit, 

177,  236 
659 
086 

626 

Building,  buildings, 

457 

air  velocities  in. 

331 

260 

classification  for  district  heating, 

651 

70 

construction,  heat  transmission  of, 

93,97 

system,     511,539,547 
ipparatus,                186 
188 

district  heating, 
fuel  requirements  of 
heat  capacity  of, 

644 
470,  482 
481 

189,  190,  194 

hot  water  supply  to, 
intermittently  cooled, 

611,016 
149 

73 

intermittently  heated,                139,  385, 

481,  672 

463 

materials,  heat  transmission  of, 

94 

301 

noise  in, 

301 

239  (see  also  Control1) 
ipment,                    457 
380 

saving  of  steam  in, 
tall,  infiltration  in, 
water  supply  to, 

645 
126 
599 

Burner,  burners, 

automatic  equipment, 

457 

coal. 

457 

sam  flow,               527 

conversion, 

473,  475 

377,  686 

gas, 

471 

71 

oil, 

469 

87 

By-  pass  method, 

156,  686 

675 

675 

C 

185 

Cabinets,                                       ($e$  Enclosures) 

177 

Calorie, 

686 

542,  583 

Calorific  values, 

aperature,                659 

coal, 

444 

686 

gas, 

453 

28  1  (see  also  Pans) 
284 

oil, 
Carbon  dioxide, 

451 
188 

361 

concentration  in  air, 

51 

52 

as  corrosion  agent, 

595 

ilftfi   RRft 

as  an  index  on 

wVt),  ODD 

410,  704 

combustion, 
draft  loss, 

445 
435 

Je,                           410 
377,  686 
405 

odors, 
measurement  of. 

33 

(181 

419 

Carbon  monoxide, 

418 

in  air, 

260 

416,  543 

in  garages, 

SS 

415,  470 
408 

poisoning, 
produced  by  oil  burners. 

261 
404 

406,470 

Cattle,  heat  and  moisture  produced  by, 

87 

710 


ALPHABETICAL  INDEX  TO  TECHNICAL  DATA  SECTION 


Page 

Ceilings,  heat  transmission,  107 

Central  air  conditioning  systems,  155, 197 

automatic  control  of,  244,  252 

design  of,  157 

location  of  apparatus,  158 

ratings  of,  162 

spray  type,  156 

Central  fan  heating  systems,  359,  686 

computations  for,  366 

design  of,  363 

electrical,  669 

heating  requirements  of,  362 

Charles'  law,  7 

Chart, 

of  air  densities,  J 

comfort,  44, 51 

effective  temperature,  .    40 

psychrometric,  _  t                  19«  insert 

Ringelmann,  of  smoke  densities,  682 

thermometric,  40 

Chimney,  chimneys,  423 

areas  of,  430,440 

characteristics,  425 

construction  of,  438 


Combustion, 

air  required  for, 
of  different  coals, 
of  gas, 
of  oil, 
smokeless, 

with  various  stokers, 
Comfort, 
chart, 

effective  temperature, 
heating  for, 
level, 
line, 

for  men  working, 
optimum  air  conditions  for, 
school  children, 


effect, 
efficiency  of, 
performance, 
sizes, 
Venturi, 
Cinder,  cinders, 
catching  devices, 
disposal  of, 


78,81,119,126,517,686 
426, 480 
429 

430, 440 
425 
260 
265 
267 


Circular  equivalents  of  rectangular  ducts,  333 

Circulator,  472 

Classrooms,  (see  Schools) 
Cleaners,  air                     (see  A  »>,  Cleaning  Dt, vices) 

Clearance,  window  sash,  122 
Coal,                   (see  also  Anthracite;  Coke;  Lignite) 

analysis  of,  443 

bituminous,  448 

calorific  value,  444 

classification  of,  443 

dust,  disposal  of,  267 

dustless,  450 

pulverized,  450 

semi-bituminous,  449 

size  of,  '               445 
Coal  burning  systems, 

automatic  control  of,  247 

automatic  firing  equipment,  457 

boilers,  405 

combustion  rate,  376 

draft  required  for,  434 

estimates  of  heating  costs,  488 

fuel  requirements,  calculation,  479, 485 

furnace  requirements,  370, 383 

hand-fired,  450 

stokers,  457 
Codes,  , 

A.S.H.V.E.  codes  and  standards,  704 
for  grinding,  polishing,  and  buffing  wheels,  348 
for  proportioning  warm  air  heating  plants,  471 

for  use  of  refrigerants,  157 
Coefficients  of  heat  transmission, 

(see  Heat  Transmission,  Coefficients} 
Coils, 

booster,  owl 

cooling,  211, 379 

evaporator,  211 

heating,  213, 301 

hot  water,  204 

pipe,  549 

radiator,  491 

steam,  20$ 

Coke,  444 

combustion  of,  450 

Cold,  effects  on  human  body,  36 

Collectors,  dust,  354 

Combined  system, 

air  conditioning  equipment,  229,  318 

automatic  controLof,  246 

central  f  an,  %$® 


Page 
443 

445,  454 
447 
453 

467,  470 

409 

457 

38 

44,  46,  51 

'      39 

657 

495 

43, 46,  48,  687 
46 

42,46 
46 


zone,  43  (see  also  Zone,  Comfort} 

Compliance,  of  sound  insulating  materials,         310 

Compressed  air,  73 

Compressors,  211 

reversed  refrigeration,  671 

types  of,  175 
Condensation, 

on  building  surfaces,  139 

meters,  649 

prevention  of,  141,  632 

rate  in  radiators,  495 

return  pumps,  518 

in  steam  heating  systems,  503,  527 

in  winter,  46 

Condenser,  202 

design  data,  178, 188,  211 

turbine,  186, 190 

water  temperatures,  189 
Conditioning  and  drying, 

67  (see  also  Air  Conditioning) 

Conductance,  92, 113,  687 

of  air  spaces,  94 

of  building  materials,  95 

of  insulation,  95,  624 

surface,  694 

Conduction,  491, 687 

Conductivity,  92, 113,  687 

Conduit,  641 

Connections, 

for  boilers,  416, 543, 545 

branch,  505 

for  central  fan  systems,  549 

for  chimneys,  440 

for  convectors,  549 

for  district  heating  systems,  644 

for  drains,  416 

Hartford  return,  543 

for  hot  water  systems,  576 

for  indirect  heating  units,  551 

for  pipe  coils,  549 

for  radiators,  547, 572 
Construction  code  for  low  pressure  boilers,  416 
Control,  controls, 

accessory  automatic  apparatus,  242 

of  air  conditioning  equipment,  239, 250 

combined  system,  246 

split  system,  244 
automatic,                                 213, 239, 243, 250 

connecting  apparatus,  242 

of  cooling  units,  260 

of  electrical  heating,  '                672 

of  fan  motors,  295 

manual,  81, 213 

of  mechanical  warm  air  systems,  3S3 

of  natural  ventilation,  81 

of  oil  burning  equipment,  465 

of  sound,  299 

of  ateara  heating  systems,  njn  jrtw  517 

of  temperature,  243, 465, 672 

of  Unit  conditioners*  213 

«.  ,472, 657, 668, 687 

Convectora,  491,496,687 

connections  for,  649 

design  of,  497 

gravity,  491 

hemt  emission  by,  491, 498 


711 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


Convectors  (continued),  Page 

heating  capacity,  498 

performance  characteristics,  497 

Conversion  equations,  699 

Coolers, 

surface,  162 

types  of,  177 

unit,  234, 695 

Cooling,  14, 19 

with  central  fan  heating  systems,  379 

effective  temperatures  for,  50 

equipment,  design  of,  188 
evaporative,                                    72, 155,  165, 184 

of  fluids,  189 

of  human  body,  •     35 

load,  145, 214 

with  mechanical  warm  air  systems,  387 
methods,                                              155,165,202 

ponds,  190 

relative  humidities  for,  50 
towers.                                         174, 179, 187, 190 

water,  188 

Copper  pipe,  580 

Corrosion, 

of  boilers,  419 

of  industrial  exhaust  systems,  356 

inhibitors,  597 

of  pipe,  595 

tester,  596 

Costs, 

comparative,  of  fuels,  488 

of  district  heating  service,  652 

of  electrical  heating,  673 

of  unit  conditioners,  215 

Crack,  window,  122 

Cyclone  dust  collector,  354 


Dal  ton,  law  of  partial  pressures,  1 
Damper,  dampers, 

apparatus  which  operates,  242, 244, 246 

control,  80 

in  duct  systems,  881 

types  of,  381 

with  unit  ventilators,  228 

Decibel,  299, 687 

Definitions,  <685 

Degree-day,  687 

industrial,  485 

method  of  estimating  fuel  consumption,  483 

records  for  cities,  484 

Dehumidification,  14, 183 

effective  temperatures  for,  50 

methods  of,  155, 166, 202 

relative  humidities  for,  50 

Dehumldlfier,  dehumidifiers,  166,  688 

alumina,  169 

in  central  air  conditioning  systems,  156 

in  industrial  air  conditioning,  72 

silica  gel,  168 

types  of,  74 

Density,  2,688 

of  air,  6 

of  saturated  vapor,  10 

aped  fie,  3 

of  water,  20 

Design  temperature, 

dry-bulb,  147 

wet-bulb,  H7, 189 

Dew  point,  688 

relation  to  relative  humidity,  9 

temperature*  2, 251 

Diameter,  circular  equivalents  of  rectangular 

ducts,  834 

Dichlorodifluoromethane,  175, 188 

Diesel  engine,  190 

Dirt  pockets,  655 

Disc  fans,  test  code  for,  284 
Distribution  of  air,     317  (see  also  A  ir,  Distribution} 

District  heating,  639 

DIverter,  back  draft,  439 


Page 

Domestic  supply, 

hot  water,  468,  617,  671 

water,  599 

Doors, 

air  leakage  through,  122 

coefficients  of  transmission  of,  113 

natural  ventilation  through,  80 

Down-feed  piping  systems,  506,  508,  510,  513 

Downward  system  of  air  distribution,  321 


Draft, 

available, 
back,  diverter, 
calculations, 
capacity, 


intensity  required, 


in  chimneys, 
through  fuel  bed, 

mechanical, 

natural, 

theoretical, 

towers, 

Drain  connections, 
Draw-through  heating  units,    * 
Drawings,  symbols  for, 
Dripping  of  steam  pipes, 

Dry-bulb  temperature,  (see  Temperature,  Dry-bulb} 
Dry  return,  688 

Drying,  67  (sec  also  Regain) 


423 

425,  432 
439 
423 
425 
676 
688 

434, 447 
432 
436 
433 
425 
423 
425 
192 
416 
361 
701 
555 


of  lumber, 
Duct,  ducts, 

air, 

design  of, 

equal  friction  method, 
velocity  method, 

for  air  distribution, 

air  velocities  in, 

circular  equivalents, 

construction  details, 

design  of  duct  systems, 

humidity  measurement  In, 

noise  transmission  through, 

pressure  loss  in, 

for  recirculated  air, 

sheet  metal  for, 

sizes  of, 

temperature  loss  in, 

temperature  measurement  in, 

velocity  measurement  in, 
Dust, 

catching  devices, 

collectors, 

concentration  in  air, 

counter, 

disposal  of, 

industrial  exhaust  systems, 

measurement  of, 


71 

325 

325,  330 
331,333 

331 

317 


342 

330,  340,  380,  393 
681 
311 
320 
393 
352 

320,333,347 

365,  391 

677 

678 

40,250,688 
265 
354 

271,081 
681 
267 
345 
681 


Dynamic  equilibrium,  Carrier's  equation  for, 

E 

EDR,  equivalent  direct  radiation,  694 

Effective  temperature,   (M«  Tempsrature,  Rffetilm) 

Elbow,  elbows, 

design  of,  380, 590 

equivalents,  503 

loss  of  pressure  in,  326 

resistance  in,  354 

sheet  metal  used  in,  352 

welding  of,  591 

Electric,  electrical, 

central  fan  heating  systems,  009 

current,  as  corrosion  agent,  595 

heat  equivalents,  673, 700 

heaters,  capacity  of,  673 

heating,  007 

auxiliary,  072 

of  hot  water,  071 

heating  elements,  205»  697,  000 

with  unit  heaters,  220 

lamp  bulbs,  heat  from,  139 


712 


ALPHABETICAL  INDEX  TO  TECHNICAL  DATA  SECTION 

Page 

Page 

Eliminator  plates  and  baffles, 

183,  187 

Fatigue,  human,                                                       37 

Emissivity, 
Enclosures, 

660 

Filter,  filters,                                                 212,  267 
automatic,                                                             275 

concealed  heaters, 

497 

cloth,                                                                      354 

convectors, 

497 

design  of,                                                     212,  274 

effect  of, 

495 

dry  air,                                                                27& 

unit  air  conditioners, 

210 

installation  of,                                                     277 

Engines, 

resistance  of,                                                       379 

Diesel, 

190 

for  sound,                  *                                          311 

internal  combustion, 
Enthalpy, 
Entropy, 

189 
18,  689 
23,  689 

unit  type,                                                              274 
viscous  type,                                                       274 
Fire  walls,                                                               352 

Equations,  conversion, 
Equilibrium, 

699 

Fittings,                   579  (see  also  Connections;  Pipe) 
areas  of,                                                                 587 

dynamic, 

2 

flanged,                                                                    589 

hygroscopic, 

70 

lift,                                                                         514 

Equipment  room,  design  of, 

307 

screwed,                                                               587 

Equivalent,  equivalents, 

welding,                                                       592,  594 

circular, 

333 

Flame,  with  oil  burners,                                 464,  469 

direct  radiation, 

138,  559,  694 

Flanges,  welding  neck,                                    592,  594 

elbow, 
evaporation, 
heat, 

563 
411,  689 
699 

Floors,  heat  transmission  through,                       107 
Flowers,  temperatures  for  greenhouses,                 72 

of  'air  infiltration, 

128 

Fluid,  fluids, 

of  brake  horsepower, 
electrical, 

138 
673,  700 

cooling  of,                                                           189 
formula  for  flow  of,                                            325 

mechanical, 

692 

meters,                                                              648 

length  of  run, 

351 

Foodstuffs, 

square  feet, 

492 

regain  of  moisture  of,                                         66 

Ethylene  gas,  in  ripening  bananas 

3,                                   71 

temperatures  and  humidities  for  processing,     68 

Eupatheoscope, 

663,  672,  683 

Force,                                                                   689 

Evaporation, 

26,  28,  72,  193 

Forge  shops,  heat  given  off  in,                              83 

equivalent, 

411,689 

Formulae,  conversion,                                          699 

from  human  body, 

35.  59 

Foundries,  heat  given  off  in,                                  83 

rate  of, 
from  water  pans, 
Evaporative  cooling, 
Evaporators, 

26 

496 
72,155,165,184 
211 

Freezing, 
of  cooling  water,                                                 193 
insulation  against,                                       •     631 

Exfiltration, 

119,303,479 

Friction, 

Exhaust  systems, 
industrial, 
Expansion, 
of  joints, 
of  pipe, 
in  steam  piping, 

345 
346,  351 

542,  641 
542,581,641 
541 

in  chimneys,                                                   428 
coefficients,                                                   329 
heads  in  pipes,                                                  562 
in  heating  units,                                               362 
losses  in  ducts,                                   327,  329,  333 
in  water  pipes,                                                606 

tanks, 

574 

Fuel,  fuels,        443  (see  also  Anthracite;  Coal;  Coke; 

Exposure  factors, 

135 

Gas;  Lignite;  Oil} 

bed,  draft  loss  through,                                 433 

F 

burning  equipment,  automatic,                       457 

Fan,  fans, 

281 

comparative  heating  costs,                              488 
requirements,                                                  479 

as  accessory  apparatus, 
A.S.H.V.E.  test  code  for, 
attic, 
booster,  equipment, 

198,  609 
284 
233,  380 
150,  402 

degree-day  method,                                    483 
method  of  approximation,                          485 
saving  during  non-heating  periods,                 481 
utilization  of,                                                    479 

brake  horsepower, 
control  of, 
for  cooling, 
designation  of, 
drives,  arrangement  of, 

187 
291 
380 
291 
292 

Fumes,                                                          259,  689 
industrial  exhaust  systems,                             345 
toxicity  of,                                                        263 
Fundamentals  of  heating  andjair^conditioning,      1 

dynamic  efficiency  of, 
efficiency  of, 

283 
283 

Furnace,  furnaces,                                              689 
design  of,                            376,  383,  396,  408,  451 

in  electrical  heaters, 

689,  672 

door  slot  openings,                                          446 

furnaces, 

375,  387 

efficiency  of,                                                    480 

for  gas-  fired  furnaces, 
for  industrial  exhaust  systems, 

472 
355 

hand-fired,                                                      451 
performance  curves  of,                                    397 

mechanical  draft. 

425 

ratings  of,                                                      387 

mechanical  efficiency  of, 

283 

types  of,                                          375,  387,  471 

motive  power  of» 

293,  296 

volume,                                                         689 

control  of, 

295 

for  warm  air  systems,                               375,  396 

mountings, 

295 

Furnacestat,                                                    383 

operating  characteristics, 

284 

operating  velocities, 

289,  290 

G 

performance  of, 
quietness  of, 
ratings  of, 

281,289 
225 

288 

Gage*  gages, 
draft,                                                             676 

selection  of, 

287,290,293 

pressure,                                                        675 

static  efficiency  of, 

283 

Steam,                                                           416 

systems  of  heating, 
tip  speeds, 

859 

289 

vacuum,                                                       675 
Galvanometer,                                                   676 

total  efficiency  of, 

283 

Garage,  garages, 

types  of, 

281,286,289,402 

air  flow  necessary  in,                                       84 

in  unit  conditioners, 

212 

A,S.H,V,E,  ventilation  code,                            S8 

In  warm  air  systems, 

377,  886,  402 

heaters  for,                                                  473 

713 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

Page 

Heat  (continued), 

Page 

Gas,  gases, 

exchanger, 

186,  188 

calorific  value, 

139,  453 

shell  and  tube, 

177 

in  chimneys, 

428 

flow  meter, 

92,  683 

flue,  analysis, 

682 

gain, 

fuel, 

from  fixtures  and  machinery, 

153 

manufactured, 

453 

from  outside  air, 

153 

natural, 

453,  480 

to  be  removed, 

161 

454 

sources  of, 

214 

scrubbers, 

267 

infiltration  equivalent  of, 

128 

toxicity  of, 

263 

latent, 

690 

Gas-fired  appliances, 
accessory  conditioning  equipment, 
automatic  control  of, 

205,  248 
199 
247,  471 

loss, 
of  water  vapor, 
of  the  liquid, 

57 
10 
23,  691 

boilers                                                  407 

415,  471 

loss, 

selection  factors, 
carbon  monoxide  produced  by, 
chimneys  for, 
classification, 
control  of, 
conversion  burners, 
furnace  requirements  for, 
heating  costs,  estimates  of, 
installation, 
rate  of  gas  consumption, 
ratings  of, 
types  of  space  heaters, 
used  with  unit  heaters, 
warm  air  furnaces, 
Gaskets, 

474 
475 
438 
470 
472,  474 
473,  475 
376,  3S3 
488 
475 
474,  486 
475 
472 
226 
471 
590 

from  bare  pipe, 
computation  of,                         91,  140, 
determination  of,                     131,  138, 
effect  of  insulation  on, 
from  human  body,                            57, 
by  infiltration, 
by  intermittently  heated  buildings, 
latent, 
from  piping  of  gas-fired  furnaces, 
by  radiation, 
sensible, 
to  unheated  rooms, 
maximum  probable  demand, 
mechanical  equivalent  of, 
of  the  liquid, 
produced  by  cattle, 

021 
600,  072 
479,  057 
033 
657,  059 
128,  479 
385 
57 
474 
41,  657 
67,  659 
116 
131 
692 
23,  691 
87 

Glass, 
heat  transmitted  through, 

113,  152 

produced  by  human  body, 
pump, 

34 

671 

solar  radiation  through, 

149 

radiation, 

693 

window,  area, 

92 

regulation  in  man, 

34,  38 

Glossary  of  terms, 

685 

sensible, 

165,  690 

Grates, 

690 

of  air, 

10 

areas  of,                                371,383,396, 

414,  690 

loss, 

57,  659 

of  furnaces,                                 371,383, 

396,414 

solar, 

146 

of  stokers, 

457 

sources  of, 

657,  670 

Gravity 

other  than  heating  plant,               138, 

170,  482 

491 

total, 

15,  251 

gravity-indirect  heating  systems, 

499 

of  saturated  steam, 
transmission, 

23 
91,  657 

specific,                                                              — 
steam  heating  systems,                                      503 
one-pipe  air-vent,                    503,  500,  535,  547 
two-pipe  air-vent,                           507,  536,  547 
warm  air  heating  systems,  design  of,              389 
Greenhouses,  temperatures  for,                             72 
Grille,  grilles,                        690  (see  also  Rt&isters) 
anemometer  readings  through,                       679 

through  air  spaces, 
through  building  materials, 
calculations, 
coefficients, 
of  ceilings, 
combined, 
of  doors, 
of  floors  i 

94 
94 
91 
91,114 
107 
115 
113 
107 

for  concealed  heaters, 

498 

95 

of  roof  ventilators, 

80 

of  roofs, 

110 

for  warm  air  systems, 

317,  393 

of  skylights, 

113 

of  walls, 

100 

H 

of  windows, 

113,  152 

convection  equation, 

658 

Hartford  return  connection, 
Health,  effect  of  air  pollution  on, 

505,  545 
260 

definition  of  terms  used, 
effects  of  solar  radiation  on, 

92 
148 

690 

formulae  » 

02 

absorbed  by  building:  structure, 
air  infiltration  equivalent  of, 
capacity,                                           60, 
of  buildings, 
of  leader  pipes, 

139 
128 
149,  690 
481 
391 

through  glass, 
measurement  of, 
in  surface  coolers, 
by  surfaces  not  exposed  to  the  sun, 
symbols  used  in  formulae, 

to  IKl/ke 

113,  152 
683 
103 
146 
92 
c>i 

conduction, 

687 

13.O1CS, 

time  lag, 

iff  A 

149 

of  air  and  water  vapor, 
of  dry  air, 
of  saturated  water  vapor, 
convection, 
conversion  equations, 
demand,  factors  governing, 
effects  on  human  body, 
electrical  equivalents  of, 
emission, 

10 
15 
18 
687 
699 
131 
35 
673,  700 

Heaters, 
for  domestic  hot  water,                                 616 
electric,                                                        86S 
capacity  of,                                                673 
radiant,                                                  472,  672 
space,                                                          472 
unit,                 .                                         219,  695 
wall,                                                              472 
Heatinft,     '                                      (set  aha  Haai} 

of  convectors. 

491,498 

costs,  relative. 

488 

by  radiation, 

660 

district, 

264,630 

of  radiators, 

491,498 

effect  of  radiators, 

408 

equivalent,  equivalents, 
of  air  infiltration. 

699 
128 

electrical, 
elements,  electric, 

$67 
OSS 

of  brake  horsepower, 

138 

fundamentals  of, 

1 

electrical, 

673,  700 

load. 

m 

mechanical, 

692 

medium, 

600 

714 


ALPHABETICAL  INDEX  TO  TECHNICAL  DATA  SECTION 


Heating  (continued}, 
radiant, 

by  reversed  refrigeration, 
surface, 

square  foot  of, 
systems, 
district, 
electrical, 
fan, 

gravity  warm  air  furnace, 
hot  water, 
mechanical  warm  air  furnace, 
radiant, 
steam, 
units, 

blow-through, 
central  fan, 
draw-through, 
Henry  and  Dalton,  law  of, 
Hoods, 

axial  velocity  formula, 
canopy, 
design  of, 

for  exhaust  systems, 
of  furnaces, 
Horsepower, 
boiler, 
brake, 
fan, 

heat  equivalent  of, 
Hot  box, 
Hot  plate, 

Hot  water  heating  systems, 
forced  circulation, 
gravity  circulation, 
installation  of, 
mechanical  circulation, 
Hot  water  piping, 
Hotels, 

stokers  suitable  for, 
temperatures  of  in  winter, 
water  supply, 
Humidincation, 
atomizatlon  for, 
effective  temperatures  for, 
methods  of, 
relative  humidities  for, 
for  residences, 
systems  of, 
with  water  pans, 
in  winter, 

Humidifier,  humidifiers , 
with  fan  systems, 
.  types  of  f 

with  uriit  air  conditioners, 
Humidistat, 
Humidity, 
absolute, 

effect  of  on  acoustics, 
for  industrial  processing, 
measurement  of, 
relative,          8,  691  (see  aU 
A.S.H,V*B.  standards, 
in  comfort  zone» 
effect  on  moisture  regain, 
relation  to  dew  point, 
specific, 
Hygrodeik, 

Hygroscopic  materials, 
with  duatleas  coal, 
moisture  content, 
proteasing  of, 
regain, 
Hygrostat, 


Ice.  in  air  conditioning, 
Inch  of  water, 
Industrial, 
air  conditioning, 
apparatus  for, 

automatic  control  of, 
air  pollution, 


Page 

Industrial  (continued), 

Page 

657,  668 

cooling  systems, 

167 

206 

degree-day, 

485 

409 

drying, 

67,  290 

694 

electrical  heating  systems, 

671 

exhaust  systems, 

345 

639 

heat  sources, 

138 

667 

plants, 

65 

359 

proceSvSing  of  hygroscopic  materials, 

67 

389 

temperatures  and  humidities  for, 

68 

559 

unit  heaters, 

226 

375 

Infants,  premature, 

43,45 

657 

Infiltration, 

6S5 

503 

average, 

126 

fuel  utilization, 

479 

361 

heat  equivalent, 

128 

361,  669 

through  shingles, 

121 

361 

through  walls, 

120 

595 

through  windows, 

123 

Institutions,  water  supply  to, 

601 

349 

Instruments, 

675 

350 
349 

Insulation, 

691 

345 

asbestos  type, 

377 

corrugated, 

626,  627 

laminated, 

628,  629 

691 

of  boilers, 

420 

411,430,686 

bright  metal  foil, 

113 

characteristics  of, 

95,  625 

187 

effect  on  heat  loss, 

633 

138 
92,  683 
683 

with  electrical  heating, 
heat  transmission  through, 
for  low  temperatures, 

667 
95,  625 

624 

559 

magnesia  type, 

625 

561 
569,  573 

575 
573 

of  piping,                                                            621 
to  prevent  condensation,                    139,  141,  632 
to  prevent  freezing,                                           631 
reflective  type,                                                113 

559 

rock  wool  type, 

630 

of  sound, 

304 

460 

tables. 

95,  625 

132 

thickness  needed, 

634 

599 

types  of, 

96 

14,  183 

underground, 

635 

73 

of  vibration, 

308 

50 

Internal  combustion  engines, 

189 

202 

lonization  of  air, 

,    56 

50 

Isobaric, 

691 

252,  496 

Isothermal, 

7,691 

194 

496 

157,  496 

J 

309 

Joints,  expansion, 

542,  641 

72 

212 

K 

241,383,691 

7t  658,  691 

Kata  thermometer, 

51,  679 

7,691 

Kiln  drying  of  lumber, 

72 

312 

68 

680 

L 

Relative  Humidity} 

Latent  heat, 

losa, 

690 
67 

45 

of  water  vapor, 

10 

9 

Lead  poisoning, 

262 

8,691 
681 

Leader  pipes, 
heat  carrying  capacity  of, 
size  of, 

389 
391 
391,401 

67 
450 

Leakage  of  air,                  119  (see  also 

Infiltration) 

65,66 

Length  of  run,  equivalent, 

531 

67 

Libraries,  air  conditioning  of, 

69 

65,66 

Lignite, 

444 

241,  691 

Liquid,  heat  of  the, 

23,  691 

Load, 

cooling, 

60,  145 

178,204,379,387 

AQl 

design, 
heating, 

410,  689 
131 

UUJ, 

hot  water  supply, 

413 

05 
72 

radiation, 

412,  68fl 
412 

250 

Louver  fences, 

ISO 

301 

Lumber,  drying  of, 

71 

716 


AMERICAN  SOCIETY  of  HEATING  and 

VENTILATING  ENGINEERS  GUIDE,  1935 

M 

Page 

Oil,  oils  (continued),                                           Page 

Machinery, 

air  supply  for,                                         463,  467 

as  heat  source, 

138,  153 

automatic,                                                     469 

sound  insulation  of, 

307,  309 

boilers,                                               407,  409,  466 

Magnesia  insulation, 

625 

carbon  monoxide  produced  by,                      464 

Manometer, 
Masonry  materials,  heat  transmission 

678,  69  1 
through,    95 

for  commercial  use,                                          469 
control  of,                                                  247,  465 
for  domestic  hot  water  supply,                      468 

Mass, 

692 

for  domestic  use,  classification,                     462 

Mb, 

559,692 

estimates  of  heating  costs,                             488 

Mbh, 

559,  692 

flame  with,                                                   .  464 

Mean  radiant  temperature, 

657,  660 

furnace  requirements,                             376,  383 

Mechanical, 

installation,                                                       467 

draft  towers, 
equivalent  of  heat, 
warm  air  furnace  systems, 

192 
692 
375 

maintenance,                                                   468 
oil  consumption.                              469,  479,  485 
operation,                                                464,  469 

Metabolism, 

35,56 

Orsat  test,                                                     467 
specifications,                                                451 

Meters, 

calorific  value  of,                                             451 

choice  of, 

648 

classifications,                                                     451 

condensation, 

649 

as  corrosion  inhibitor,                                        597 

fluid, 

648 

cost  of,                                                                 453 

Nicholls  heat  flow, 

683 

heated  electrically,                                          671 

steam  flow, 

650 

ignition  of,                                                          452 

types  of, 

648 

preheating,                                                        469 

Methyl  chloride, 

188 

specifications,                                                    451 

Metric  units, 

701 

One-pipe  steam  heating  systems, 

Micron, 

692 

gravity  air-vent,                          503,  506,  535,  547 

7fl 

down-feed,                                                     506 

Mildew, 

{  U 

up-feed,                                                          504 

Mixture,  air  and  water  vapor, 

10 

vapor,      '                                                        508 

Moisture, 

632 

Openings, 

content, 

air  inlet,                                                              80 

of  air,                          .      .           5 

2,  65,  72,  165 

monitor,                                                                 80 

as  index  of  air  distribution, 

51 

for  natural  ventilation, 

of  hygroscopic  materials, 

66 

location  of,                                                      83 

loss  by  human  body, 

57,59 

resistance  offered  to  flow,                               84 

from  outside  air, 

153 

size  of,                                                             77 

produced  by  cattle, 

87 

types  of,                                                         78 

regain, 

65 

Ori6ce,  orifices, 

Mol, 

692 

friction  heads,                                                 568 

Monitor  openings, 

80 

steam  heating  systems,              510,  529,  539,  547 

Motive  power, 

281 

Orsat  test  apparatus,                                   467,  682 

Motors, 

293 

Outlets,                            (see  also  Registers;  Grilles) 

classification  of, 

294 

design  and  location  of,                                     169 

for  fan  operation, 

293 

Oxygen,                                                                 595 

as  heat  source, 

138,  153 

of  unit  air  conditioners, 

210 

P 

MRT,  mean  radiant  temperature, 

657,  660 

Paint, 

effect  on  radiators,                                          493 

spray  booths,                                                   351 

temperatures  and  humidities  for  processing,      69 

Natural  draft  towers, 

192 

Partial  pressures,  Dalton's  law  of,                         I 

Natural  ventilation, 

77 

Perspiration,                                         35,  04,  58,  60 

Nicholla  heat  flow  meter, 

683 

Petterson-Palmquist  apparatus,                         681 

Noise,                                             (see 

also  Sound) 

Phon,                                                                300 

in  buildings, 
control  of, 
through  ducts, 
with  warm  air  systems, 

1*kV*»l 

301 
303 
311 
378 

Pipe,  piping,                                                     579 
bare,  heat  loss  from,                                       621 
bends,                                                              542 
capacities,                                  (set  Pipe,  Siws) 

jeVcl, 

acceptable, 
of  compressors, 
of  fans, 
of  unit  heaters, 
of  various  apparatus, 
measurement  of, 
Nozzle, 
air  spray^ 
oil  atomiser, 
water  spray, 

303 
175 
288 
225 
302 
300 
183 
139,  322 
470 
73 

conduit,                                                         041 
connections,          643,  644  (see  also  Connections) 
corrosion  of,                                                    595 
dimensions  of,         580,  821  (w*  also  Pipt>  Si&$$) 
down-feed  systems,                   506,  SOS,  510,  513 
expansion  of,                                     541,581.041 
fittings,          580  (see  also  Connections;  Fittings) 
for  water  supply,                                        604 
welding  fittings,                                   590,  621 

flexibility  of,                                                 '  581 

, 

friction, 

O 

of  air  in,                                                       327 

Odors, 

51 

heads  in,                                                   563 

of  human  origin, 
concentration, 
removed  byloutsJde  air, 

on,  oils, 

33 
54 
54 

gaskets,                                                        501 
hangers,                                                           585 
heat  loss  from,                                        474,  621 
for  hot  water  heating  systems,                559,  575 
insulation  of,                                                  621 

atomi»ation  of, 

463 

Joints,                                                     542.  641 

burner,  burners, 

leader,                                                          391 

accessory  conditioning  apparatus, 

109 

radiators,                                                     491 

air  for  combustion, 

407 

scale  in,                                                       595 

ALPHABETICAL  INDEX  TO  TECHNICAL  DATA  SECTION 


Pipe,  piping  (continued}, 
sizes, 

for  boiler  runouts, 

for  central  fan  systems, 

for  conyector  connections, 

dimensions, 

for  district  heating, 

for  domestic  hot  water, 

effects  of  variation  of, 

elbow  equivalents, 

equivalent  length  of  run, 

friction  head, 

of  orifices  in  unions, 
for  Hartford  return  connection, 
for  hot  water  heating  systems, 
forced  circulation  systems, 
gravity  circulation  systems, 
for  indirect  heating  units, 
for  pipe  coil  connections, 
for  radiator  connections, 
return,  capacity  of, 


Page       Pressure,  pressures  (continued}, 


Page 


steam, 

underground, 

tables, 

tees, 

for  underground  steam, 

for  water  supply, 

weights, 

steam,  capacity  of, 
for  steam  heating  systems, 
supports, 
sweating, 

systems,  down-feed, 
systems,  up-feed , 
tax, 

tees,  dimensions  of, 
threads, 
tunnels, 
types  of, 
underground, 

insulation  of, 

steam, 

for  unit  conditioners, 
for  unit  heaters, 
up-feed  systems, 
valves, 

water  supply, 
weights  of, 
welding, 
Pitot  tube, 


steam, 

545               in  district  heating,  640 
549                drop,                                                   505,  528, 533 

549                initial,  528 

580, 582               in  orifice  systems,  516 

640               saturated,      ,  23 

611                in  sub-atmospheric  systems,  515 

567  total,  695 

563  vapor,                                                            27, 696 
531           velocity,  696 

564  water,  26, 602 

568  Processing,  65 
543           cooling  systems,  167 
560           industrial,  temperatures  and  humidities  for,       68 
562           of  textiles,  67 

569  unit  air  conditioners,  200 
551           unit  heaters,  227 
549       Propeller  fans,  test  code  for,  284 
547        Psychro  meter,  693 
534           sling,  680 

Psychrometric, 

chart,  insert 

explanation,  19 

tests,  39 

Pump,  pumps, 

circulating,  573 

condensation  re-turn,  518 

heat,  671 

vacuum,  515 

ratings  of,  519 

Pyrheliometer,  146 

Pyrometer,  .  677, 693 

mercurial,  677 

optical,  677 

radiation,  677 

thermo-electric,  677 


528,  532,  533 
639 

533, 582 
587 
639 
602 
580 
529 

503, 527 
585 
632 

506, 508,  510,  513 

504,507,509,512 

413 

587 

585, 589 
643 
579 

635 
639 
213 
225 

504,507,509,612 
592 
599 
680 
590 
•  678 


Plastering  materials,  heat  transmission  through,  97 
Plenum., 

chamber,                                         '  692 

systems,  automatic  control  of,  246 

Plumbing  fixtures,  599 

Pollution  of  air,  259 

Ponda,  cooling,  190 

Potassium  permanganate,  191 

Potentiometer,  676, 692 

Power,  692 

conversion  equations.  699 

electric,  073 

Preclpltators,  dust,  266 

Pressure,  pressures, 

absolute,  085 

in  heating  unit,  362 

measurement  of,  678 

apparatus  sensitive  to,  241 

atmospheric,  675, 686 

for  atomizauon,  18* 

barometric,  428, 676 

basic,  8 

conversion  equations,  700 

drop,  formula  for,  640 

dynamlc'  AM  Aon 

gage,  675, 690 

loss  through  ducts,  325 

measurement  of,  675 

partial*  Daiton'a  law  of,  1 

refrigerating  plant,  188 

of  saturated  vapor,  10 

static,  371, 694 


Quality, 

of  air,  51, 55 

A.S.H.V.E.  ventilation  standards,  49 

impaired  by  recirculation,  55 
Quantity, 

of  air, 

A.S.H.V.E.  ventilation  standards,  49 

blown  by  wind,  77 

measurement  of,  678 
necessary  for  ventilation,             52, 54,  83, 159 

of  cooling  water,  188 


Radiant  heat,  472 

thermometric  chart  does  not  apply,  41 

Radiant  heating,  657 

Radiation,  693 

by  black  body,  660 

direct,  control  of,  243 

equivalent  direct,  138, 559, 694 

heat  loss  by,  41,657 

by  human  body,  38, 41,  59T  657 

as  index  of  fuel  consumption,  485 

load,  412 

with  oil  burners,  464 

by  radiators,  491 

solar,                                      .  146 

curative  value  of,  56 

effect  on  heat  transmission,  148 

occlusion  of,  261 

Radiator,  radiators,  491,693 

A.S.H.V.E.  code  for,  498 

column,  687 

condensation  rate  in,  495 

connections,  547, 572 

enclosed,           "  495 

Cared,  473 

i  emission  of ,  491,498 

heating  capacity  of,  498 

heating  effect  of,  493 

for  hot  water  systems,  569 


717 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

Radiator,  radiators  (continued), 

Page 

Resistance, 

Page 

output  of, 
paint,  effect  of, 

492 
493 

of  bright  metallic  surfaces, 
of  building  materials, 

113 
95,  632 

ratings  of, 

492 

in  ducts, 

354 

shields, 

496 

of  exhaust  systems, 

355 

types  of, 

491 

of  filters, 

379 

warm  air, 

473 

of  insulators, 

95,  632 

Rain,  as  dust  catcher, 

268 

of  sound-insulating  materials, 

308,  310 

thermal, 

694 

Ratings, 

Resistor, 

667,  671 

of  air  cleaning  devices, 

272 

Restaurants, 

of  boilers, 

410,475 

tobacco  smoke  in, 

52,212 

for  central  fan  conditioning, 

162 

water  supply  to, 

618 

of  concealed  heaters, 

497 

Return, 

of  fans, 

288 

dry, 

688 

of  furnaces, 

387 

pipe,  capacity, 

534 

of  gas-  fired  appliances, 

475 

wet, 

696 

of  noises, 

302 

Ringelmann  chart, 

682 

of  pressure  reducing  valves, 

539 

Rock  wool  insulation, 

630 

of  radiators, 

492 

Roof,  roofs, 

for  surface  coolers, 

162 

coefficients  of  transmission  of, 

110 

for  unit  air  conditioners, 

214 

conductivities  of, 

97 

of  unit  coolers, 

236 

solar  radiation  on, 

148,  150 

of  unit  heaters, 

223 

ventilator, 

80 

of  unit  ventilators, 

231 

of  vacuum  pumps, 

519 

S 

Receivers,  alternating, 

524 

Refrigerants, 

693 

Salts  in  cooling  water, 

187 

codes  for  use  of, 
types  of, 
Refrigerating  capacity, 
Refrigerating  plant, 

157 

176,211,379 
170 
186,  188 

Scale, 
in  boilers, 
centigrade, 
on  equipment, 
Fahrenheit, 

419 
676 
187 
676 

compressor, 

190 

in  pipe, 

595 

operating  methods, 

171 

Reaumur, 

676 

size  of, 

171 

School,  schools, 

steam  jet  system, 

173,  190 

air  distribution  in, 

317 

Refrigeration, 

air  flow  necessary  in, 

84 

cycle, 

167 

optimum  air  conditions, 

46 

dehumidification  by, 
diagram  of, 

106 
167 

stokers  suitable  for, 
temperature  of,  in  winter, 

460 
132 

direct  expansion  system, 

202,215 

ventilation  in, 

46,  310 

indirect  expansion  system, 

203 

Scrubbers, 

267,  273 

mechanical, 

202 

air, 

183 

reversed, 
steam  jet, 

206,  671 
204,  675 

Sensible  heat, 

105,  600 

ton  of, 

170,  605 

of  air. 

10 

ton-day  of, 
unit  coolers, 
vacuum, 

695 
234 
204 

loss, 
Settling  chamber, 
Sheet  metal,  for  ducts, 

57,  651) 
260 
352 

Shingles,  air  leakage  through, 

121 

of  hygroscopic  materials, 
standards  of  commercial, 
Registers,                               603 

66 
70 
(see  also  Grilles) 

Silica  gel, 
regain  of  moisture  of, 
system  of  adsorption, 
Silicosis, 

66 
168 
262 

with  gas-  fired  furnaces,                                  472 
with  mechanical  warm  air  furnace  systems,    381 
sizes,                               '                          ^Q<9  A'"11 

Sizes  of  pipe, 
"  Skylights, 

(m  Pipe,  Sims) 
80,  113 

and  stacks. 

81 

Sling  psychrometer, 

680 

of  unit  air  conditioners, 

210 

Smoke, 

259,  098 

Reheatcrs 

161 

abatement  of, 

263 

measurement  of, 

082 

Relative  humidity, 

50,  691 

recorders, 

082 

apparatus  sensitive  to, 

241 

tobacco, 

52,  212 

A.S.H.V.B.  standards, 

49 

Solar  heat, 

146 

for  banana  ripening* 
in  comfort  zone, 
control  of, 

71 
45,48 
251 

Sound, 

absorption  coefficients, 

299 
803 

effect  on  sound, 
in  industrial  plants, 
of  libraries, 
for  lumber  drying, 
mildew 

812 
65 
70 
72 
70 

control, 
effect  on  duct  design, 
effect  of  humidity  on, 
effect  of  temperature  on, 
insulation  of, 

299 
341 
B12 
312 
804 

for  processing, 
in  public  buildings, 
in  residences, 

68,70 
48 
496 

intensity, 
measurement  of, 
in  steam  heating  systems, 

399 

aoo 

580 

for  textile  testing 

70 

Specific  density, 

3 

from  water  pans, 

496 

Specific  gravity, 

2,003 

Research  residence, 

879,899 

of  fuel  gas, 

454 

Residences, 

Specific  heat, 

3,693 

air  distribution  in, 

317 

mean,  of  water  vapor, 

3 

humidification  of, 

252,496 

of  water, 

26 

oil  burners  for, 

462 

Specific  humidity, 

8 

conversion  burners. 

473 

Specific  volume, 

3,698 

stokers  for, 

458 

of  saturated  steam, 

m 

718 


ALPHABETICAL 

INDEX  TO  TECHNICAL  DATA  SECTION 

Page 

Storage, 

Page 

Split  system, 

694 

of  hot  water, 

617,671 

air  conditioning  equipment, 
automatic  control  of, 

229,  319 
244 

temperatures  and  humidities  for, 
Storm  sash, 

68 
122 

central  fan, 

359 

Sub-atmospheric  systems,       514,  520, 

529,539,547 

Spray, 

Suction, 

booths  for  painting, 

351 

static,  in  exhaust  systems, 

347 

cooling, 

unloaders, 

211 

ponds, 

190,  192 

Sulphur  dioxide  in  air, 

09 

efficiency  of, 

191 

Summer, 

towers, 

191,  192 

care  of  heating  boilers, 

419 

distribution  of, 

73 

comfort  zone, 

43,46,48 

generation  of, 
type  of  central  station  system, 

73 

156 

conditioning,  apparatus  for, 
desirable  indoor  conditions  in, 

155,  159 
48,50 

water  coolers, 

177 

temperatures, 

147 

Square  foot  of  heating  surface, 

604 

wind  velocities  and  directions, 

147 

Stack,  stacks, 

81,389 

Sun 

effect, 

7S 

diurnal  movement  of, 

148 

height, 

694 

effect, 

60 

size  of, 
system  with  registers, 
wall, 

83,392,401 
SI 
392 

effect  on  heating  requirements, 
factor  of  cooling  load, 

517 
152 

Stairways, 

127 

cooling, 

156 

Standards, 
air  conditioning, 
A.S.H.V.E,  codes  and  standards. 

33,48 
48,  704 

equipment, 
air  conditioning, 

162 
156 
162 

commercial  regain, 

70 

extended, 

497 

for  pipe, 
for  ventilating  industrial  plants, 

581 
262 

gravity-indirect  heating  systems, 
heating, 

499 
689,  691 

for  welding, 

f>91 

square  foot  of, 

694 

Steam, 

188,  694 

radiant  heating, 

661 

205 

secondary  , 

466 

condensing  rates, 
flow,  Babcock's  formula, 
heat  content  of, 
as  heat  source, 

494 
527 
18 
670 

Symbols, 
for  drawings, 
for  heat  transmission  formulae, 
Synthetic  air  chart, 

701 
92 
694 

heating  systems, 

503,  670 

air-vent,                                    503, 

507,  535,  536 

T 

atmospheric, 

511,539,547 

classification, 

503 

Tank,  tanks, 

condensation  return  pumps, 

518 

for  domestic  water  supply, 

605 

connections,               (see  Connects 

5ns;  Fittings') 

expansion, 

574 

corrosion  of, 

595 

flush, 

605 

design  of, 

511,  527 

Tees,  dimensions  of, 

587 

clirt  pockets, 
district  heating, 
dripping  of, 
electric 

555 
639 
555 
670 

Temperature, 

absolute, 
of  air  leaving  outlets, 

685 
159,  364 

equivalent  length  of  run, 
gravity  systems, 
one-pipe,                             503, 

531 
503 
500,  535,  547 

ammonia, 
apparatus  sensitive  to, 
for  banana  ripening, 

188 
239,  676 

71 

two-pipe, 

507,  536,  547 

of  barns, 

87 

with  high-pressure  steam,                           539 
mechanical,                                                    503 
oriflce,                                     516,  529,  539,  547 
pipe,                                    527  (see  also  Pipe) 
capacity,                                          529,  532 
sizes,                                                &2S.  532 

basic, 
body, 
changes,  effect  on  human  beings, 
of  chimney  gases, 
in  cities, 
of  city  water. 

34,  30,  658 
37 
428 
136,  147,  484 
179 

pressure  drop  i», 
sub-atmospheric,             514,  520, 
type?  of, 
vacuum,                   513,519,522, 

505,  515 
529,  539,  547 
503 
538,  547,  696 

commonly  specified,     * 
control  of, 
of  cooling  water, 
dew-point, 

3 

243,  465,  672 
188 
2,251 

vapor, 
one-pipe, 
two-pipe, 
water  hammer  in, 
high  pressure, 
jet  apparatus, 
meters  for, 
pressure, 
properties  of, 
requirements  of  buildings, 
saturated,  properties  of, 
savings  in  use  of, 
tables, 

529,  547 
508 
508,  537 
530 
539 
204,  675 
648 
539 
22,  670 
651 
28 
645 
7,  18,  28 

difference, 
between  floor  and  ceiling, 
desired,  determination  of, 
in  stacks  and  leaders, 
dry-bulb, 
as  index  of  air  distribution, 
maximum  design, 
specified  in  winter, 
for  drying  lumber, 
efject  on  moisture  regain, 
efcrect  on  sound, 
effective,                            41,  49 
A.S.H.V.E.  standards, 
chart, 

134,  494 
83 
81,  301 
1,  9,  49,  688 
51 
147 
132 
72 
67 

,  133,  165,  688 
49 
40 

trap, 
underground, 
in  unit  heaters, 

039 
221 

for  maximum  comfort, 
optimum, 
scale. 

48,  659 
43 
39,  50 

Stokers, 

of  gas  flame, 

automatic  control  of, 

247,  383 

for  greenhouses, 

72 

design  of, 

457 

in  industrial  plants, 

65,  485 

economy  of, 
mechanical, 
type*  of, 

457 
457 
408,  457 

in  industrial  processing, 
inside,                                        48 
surfaces, 

68,70 
,  133,  364,  483 
860 

719 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


Temperature  (continued)*  Page 

low,  insulation  for,  624 

of  mean  interior  surface,  660 

mean  radiant,  657,  660 

measurement  of,  87,  239,  676 

in  occupied  space,  48,  52,  483 

outside,  134, 146,  483 

radiation-convection,  663 

range  of  cooling  equipment,  189\ 

records  of  cities,  136, 147, 484 

at  registers,  391,  396,  500 

room,           .  9, 195, 483, 677 

sensations,  38 
surface, 

of  man,  658 

mean  interior,  660 

systems  for  control  of,  243, 465,  672 

for  textile  testing,  70 

thermo-equivalent  conditions,  39 

value  used  in  calculations,  124,  364 

wet-bulb,  2, 15,  680,  696 

average,  189 

design,  147, 189 

as  index  of  air  distribution,  51 

maximum,  189 

Terminology,  685 

Test  methods,  675 

Textile,  textiles, 
fibres, 

regain  of  moisture,  65 

weaving  of,  67 

temperatures  and  humidities  for  processing,     69 

testing,  standard  atmosphere  for,  70 

Theaters, 

air  distribution  in,  321 

*  cooling  in,  320 

heat  sources  in,  138 

temperatures  of,  48, 132 

Thermocouples,  676 

Thermodynamics,  694 

of  air  conditioning,  1 

laws  of,  691 

Thermo-equivalent  conditions,  39 

Thermometer, 

duct, 

globe, 

Kata, 

mercurial, 

recording, 

resistance, 

Thermometric  chart. 
Thermopile, 
Thermostat,  thermostats, 

differential, 

with  gas-fired  furnaces, 

location  of, 

with  oil  burners, 

with  radiant  heaters, 

types  of, 
Tobacco  smoke, 
Ton  of  refrigeration. 
Ton-day  of  refrigeration, 
Towers,  cooling, 

atmospheric, 

mechanical  draft, 

natural  draft, 


Traps, 

dust  catchers, 

return,  automatic, 

with  steam  heating  systems, 

types  of, 
Tube, 

Bourdon, 

Pitot, 

ihell  and  tube  heat  exchanger, 
Tuning  fork, 
Tunnels,  for  steam  pipe, 
Turbines,  with  unit  heaters, 


677 
663 
079 
676 
677 
676 
40 
676 
696 
9,262 
472 

243,  247,  383 
464,  408 
672 

81,228,239 
52, 212 
170,  695 
695 

174, 179, 187, 190 
192 
195 
192 
191 
f 

266 
524 

309, 520 
521 

675 
678 
177 

aoi 

643 
226 


Two-pipe  stearn  heating  systems,          .  Page 

gravity  air-vent,  507,  536,  547 

down-feed,  508 

up-feed,  507 

vapor,  508, 537 

down-feed,  510 

U 

Ultra-violet  light,  56,  261 

Underwriters'  loop,  505,  545 
Unit  air  conditioners,                             197,  319,  695 

accessory  apparatus,  198 

advantages,  200 

classification,  201 

costs,  initial  and  operating,  215 

design  of,  207 

functions  of,  201 

installation  of,  213 

location  of,  206 

ratings  of,                        .  214 

required  capacity  of,  ^14 

types  of,  205 

uses  of,  200 
Unit  conditioning  systems, 

197  (set  Unit  Air  Conditioners) 

Unit  coolers,  219,  695 

design  of,  234 

ratings  of,  236 

Unit  heaters,  219 

blow-through  type,  capacity  of,  220 

boiler  capacity,  224 

control  of,  243 

design  of,  219 

draw-through  type,  capacity  of,  222 

electric,  669 

output  of,  223 

ratings  of,  223 

types  of,  219,  669 

used  in  industry,  226 

Unit  ventilators,  219,  227 

capacity  of,  230 

control  of,  244 

design  of,  227 

ratings  of,  231 

Unwin  pressure  drop  formula,  640 

Up-feed  piping  systems,  504,  507,  509,  512 

Upward  system  of  air  distribution,  321 


Vacuum  pumps,  515 

Vacuum  refrigeration,  204 

Vacuum  system  of  steam  heating, 

513,  519, 522,  f)38,  547,  090 

Valve,  valves,  592 

apparatus  which  operates,  ,241 

on  boilers,  417, 544 
control, 

in  oil  installations,  468 

with  steam  heating  systems,  541, 593 

with  high  pressure  ateam,  539 

operator.  Ml 

pressure- reducing,  540 

ratings  of,  5B9 

for  radiators,  505, 500 

roughing-in  dimensions,  596 

sub-atmospheric  system,  515 

on  traps,  522 

types  of,  598 

for  water  supply,  604 

Vanes,                           '  322 

Vapor, 

mixture  with  air,  10 

pressure,  20 

steam  heating  systems,  508 
water,                                             3, 5, 9, 16, 18 

weight  of  saturated,  10 

Vegetables,  temperature!  for  greenhouses,  72 

Velocity,  00$ 

rA.S.H.V.E,  ventilation  standard*,  49 

in  ducts  of  buildings,  SB1, 841 

in  exhaust  systems,  347, 353 


720 


Roll  of  Membership 

AMERICAN  SOCIETY  of 
HEATING  and  VENTILATING  ENGINEERS 


1935 


Contains  Lists  of  Members 
Arranged  Alphabetically  and 
Geographically,  also  Lists  of 
Officers  and  Committees,  Past 
Officers  and  Local  Chapter 
Officers 


Corrected  to  January  1,  1935 

Published  at  the  Headquarters  of  the  Society 

51  Madison  Avenue,  New  York,  N*  Y. 


Officers  and  Council 

AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS 

51  Madison  Ave.,  New  York,  N.  Y. 

1934-35 

President C.  V.  HAYNES 

First  Vice-President JOHN  Ho  WATT 

Second  Vice-President , _ ,.G.  L.  LARSON 

Treasurer D.  S,  BOYDEN 

Secretary , A.  V.  HUTCHINSON 


Council 

C.  V.  HAYNES,  Chairman 
JOHN  Ho  WATT,   Vice-Chairman 

One  Year  Two  Years  Three  Years 

ALBERT  BUENC.KR  R.  H.  CARPENTER  M,  C.  BEMAN 

F.  E.  GlESECKE  J.  D,  CASSELt  E.    H,    GURNEY 

W,  T.  JONES  F.  C.  MC!NTOSU  0.  W.  OTT 

J.  F.  MclNTiRE  L.  W.  MOON  W.  A.  RUSSELL 

W.  E.  STARK 

Committees  of  the  Council 

Executive: "  John  Howutt,  Chairman;  J.  D*  Cassell,  E.  H.  Gurney. 
Finance:    R»  H.  Carpenter,  Chairman;  F.  C.  Mclntosh,  J,  F.  Mclntire. 
Meetings;   W.  E,  Stark,  Chairman;  F.  E.  Giesecke,  L.  W,  Moon. 
Membership:    F.  C,  Mclntosh,  Chairman;  Albert  Buenger,  W,  A.  Russell, 

Advisory  Council 

W,  T,  Jones,  Chairman;  Homer  Addams,  R,  P.  Bolton,  W,  H.  Carrier,  S.  E,  Dibble, 
W,  H.  Driicoll,  H.  P.,  Gant,  John  F,  Hale,  L.  A,  Harding,  H,  M,  Hart,  E.  Vernon 
Hill,  J.  D,  Hoffman,  S,  A.  Jellett,  D*  D,  Kimball,  S.  R.  Lewis,  Thornton  Lewis, 
J,  I.  Lyle,  J.  R,  McColl,  D.  M,  Quay,  C  L,  Riley,  F,  B,  Rowley,  F.  R.  Still  and  A,  C 
Willard. 

Cooperating  Committees 

A»S»H.V,E.  representative  on  National  Research  Council:    Prof,  F,  E,  Giesecke  (3  years). 

a 


Special  Committees 

Committee  on  Admission  and  Advancement:  A.  J.  Offner,  Chairman  (two  years};  J.  G. 
Eadie  (one  year),  and  E.  N.  Sanbern  (three  years). 

Publication  Committee:    W.  M.  Sawdon,  Chairman;  A.  P.  Kratz  and  M.  C.  Beman. 

Committee  on  Constitution  and  By-Laws:  Thornton  Lewis,  Chairman;  W.  T.  Jones  and 
O.  W.  Ott. 

Guide  Publication  Committee:  W.  L.  Fleisher,  Chairman;  John  Howatt,  G.  L.  Larson, 
S.  R.  Lewis,  E.  N.  McDonnell,  W.  M.  Sawdon  and  J.  H.  Walker.  J.  L.  Blackshaw, 
technical  assistant. 

Committee  on  Ventilation  Standards:  W.  H.  Driscoll,  Chairman,-  J,  J.  Aeberly,  L.  A. 
Harding,  D.  D.  Kiraball,  J.  R.  McColl,  C.  L.  Riley,  W.  A.  Rowe,  Perry  West  and 
A.  C.  Willard, 

Committee  on  Code  for  Testing  and  Rating  Condensation  and  Vacuum  Pumps:  John 
Howatt,  Chairman;  W.  H.  Driscoll,  L.  A.  Harding  and  F.  J.  Linsenmcyer. 

Committee  on  Code  for  Testing  and  Rating  Connectors:  R.  N.  Trane,  Chairman;  E.  H, 
Beling,  R.  F.  Connell,  M.  Dillman,  W.  Ewald,  J.  H.  Holbrook,  Hugo  Hutzel, 
A.  P.  Kratz,  M.  G.  Steele  and  O,  G.  Wendel. 


Committees— 1934 
Nominating  Committee  for  1934 


Chapters 
Cincinnati 
Cleveland 
Illinois 
Kansas  City 
Massachusetts 
Michigan 
Western  Michigan 
Minnesota 
New  York 
Western  New  York 
Ontario 

Pacific  Northwest 
Philadelphia 
Pittsburgh 
St.  Louis 

Southern  California 
Wisconsin 


Representative 
J.  J.  BRAUN 

C.  F.  EVELETH 

J.  J.  AEBERLY 

W.  A,  RtJSSELL 

J.  F.  TUTTLE 
W.  G.  BOALES 
S.  H.  DOWNS 
N.  D.  ADAMS 
H.  W.  FIELDER 

D.  J.  MAHONEY 
H.  S,  MOORE 
A.  L,  POLLARD 
W.  R,  EICHBERG 

F.  C.  MclNTOSH 

R.  J,  TENKONOHY 

L.  H.  POLDERMAN 

E.  A.  JONES 


Alternate 
H,  E.  SPROULL 
F.  A,  KITCHEN 
J,  J.  HAYES 

E.  K.  CAMPBELL 
W.  F.  GILLXNG 
TOM  BROWN 

J,  H.  VAN  ALSBURG 
C,  E.  LEWIS 
V.  J.  Cuca 
J,  J.  YAGER 

F.  E.  ELLIS 
LINCOLN  BOUILLON 
M.  F.  BLANKIN 

F,  C.  H00GHTEN 
R.  L.  GlFFORJD 

ERNEST  SZEKELY 


Committee  on  Research 

JOHN  HOWATT,  Chairman 

F.  E.  GIESECKE,   Vice- Chair  man 

DR.  A.  C.  WILLARD,  Technical  Adviser 

F.  C.  HOUGHTEN,  Director 
O.  P.  HOOD,  Ex-Officio  Member 

One  Year  Two  Years  Three  Years 

D.  E.  FRENCH  ALBERT  BUENGER  C.  A.  BOOTH 

F.  E.  GIESECKE  S.  H.  DOWNS  E.   K.   CAMPBELL 
L.  A.  HARDING  H.  N.  KITCHELL  JOHN  HOWATT 

A.  P.  KRATZ  H.  R.  LINN  A.  J.  NESBITT 

G.  L.  LARSON  PERRY  WEST  J.  H.  WALKER  ' 

Executive  Committee  Finance  Committee 

JOHN  HOWATT,  Chairman  A.  J.  NESBITT,  Chairman 

F.  E.  GIESECKE  H.  R.  LINN        N.  D.  ADAMS  G,  L.  LARSON 

E.  C.  EVANS  J.  H.  WALKER 

Technical  Advisory  Committees,  1934-1935 

Air  Conditions  and  Their  Relation  to  Living  Comfort:    C.  P.  Yaglou,  Chairman;  J.  J. 

Aeberly,  W,  L.  Fleisher,  D.  E.  French,  Dr.  R.  R.  Sayersand  Dr.  C.-E.  A.  Winslow. 
Air  Conditioning  in  Treatment  of  Diseases:    Dr.  E.  V.  Hill,  Chairman;  N.  D.  Adams, 

J.  J.  Aeberly,  Margaret  Ingels,  H.  R.  Linn  and  E.  L.  Stammer. 
Atmospheric  Dust  and  Air  Cleaning  Devices  (Including  Dust  and  Smoke):    H.  C.  Murphy, 

Chairman;  J.  J.  Bloomfield,  M.  I.  Dorfan,  Philip  Drinker,  Dr.  Leonard  Greenburg, 

S.  R.  Lewis,  T.  W.  Pangborn,  F,  B.  Rowley  and  Games  Slayter. 
Correlating  Thermal  Research:    R,  M.  Conner,  Chairman;  D.  S.  Boyden,  J.  C.  Fitts, 

H.  T.  Richardson  and  Perry  West. 
Corrosion:    J.  H.  Walker,  Chairman;  H.  F.  Bain,  E.  L.  Chappell,  W.  H.  Driscoll  and 

R.  R.  Seeber. 
Direct  and  Indirect  Radiation  with  Gravity  Air  Circulation:    H.  F.  Hutzel,  Chairman; 

A.  P.  Kratz,  H.  R.  Linn,  J.  F.  Mclntire,  J.  P.  Magos,  T.  A.  Novotney,  R.  N.  Trane 

and  G.  L.  Tuve. 
Gas  Heating  Equipment:    W.  E.  Stark,  Chairman;  Robert  Harper,  E.  A,  Jones,  Thomson 

King,  J .  F.  Mclntire,  E.  L,  Tornquist  and  H,  L,  Whitelaw. 
Heat  Requirements  of  Buildings:    D.  S.  Boyden,  Chairman;  P.  D.  Close,  W.  H.  Driscoll, 

IL  M.  Hart,  P,  E.  Holcombe,  V.  W.  Hunter,  E.  C.  Rack,  F.  B.  Rowley,  R.  J.  J. 

Tennant  and  J,  H-  Walker. 
Heat  Transfer  of  Finned  Tubes  with  Forced  Air  Circulation:    F.  B.  Rowley,  Chairman; 

H.  F.  Bain,  H.  F,  Hutzel,  W.  G.  King,  A.  P.  Kratz,  E,  J.  Lindseth,  L.  P.  Saunders, 

G,  L.  Tuve  and  W.  E,  Stark. 
Minimum  Temperature  and  Method  of  Introduction  of  Cooling  Air  in  Classrooms:    Perry 

West  Chairman;  J,  D.  Cassell,  S.  R.  Lewis,  J,  R.  McColl,  A.  J,  Nesbitt,  G,  E,  Otis 

and  C.-E,  A,  Winslow, 
Oil  Burning  Devices:    H,  F,  Tapp,  Chairman;  Elliott  Harrington,  F.  B.  Howell,  J,  H. 

Mcllvaine,  L.  E.  Seeley  and  T,  H.  Smoot. 
Pipe  and  Tubing  (Sites)  Carrying  Low  Pressure  Steam  or  Hot  Water:    S.  R.  Lewis, 

Chairman;],  C.  Fitts,  F.  E.  Giesecke,  H,  M,  Hart,  C.  A.  Hill,  R.  R.  Seeber  and 

W.  K.  Simpson. 
Refrigeration  in  Relation  to  Air  Treatment:    A.  P,  Kratx,  Chairman;  E.  A,  Brandt,  John 

Everetts,  Jr.,  E.  D.  Milener,  K.  W.  Miller,  E.  B.  Newill,  F,  G.  Sedgwick  and  J,  H, 

Walker. 
Sound  in  Relation  to  Heating  and  Ventilation:   V.  0.  Knudsen,"  Chairman;  Carl  Ashley, 

C,  A.  Booth,  F.  C.  Mclntosh,  R.  F.  Norris,  J.  S.  Parkinson,  C.  H.  Randolph,  J.  P. 

Reis  and  G.  T.  Stanton. 
Ventilation  of  Garages  and  Bus  Terminals:    E»  K.  Campbell,  Chairman;  S.  H.  Downs, 

T.  M.  Dugan,  E»  C.  Evans,  F.  H.  Hecht,  H,  L,  Moore  and  A.  H.  Sluss. 


Officers  of  Local  Chapters 

1934-35 


Cleveland 

Headquarters,  Cleveland,  Ohio 

Meets:  Second  Thursday  in  Alonth 
President,  M.  F,  RATHER 

2142  East  19th  Street 
Secretary,  E.  J.  VERMERE 

2125  Wyandotte  Avenue 

Cincinnati 

Headquarters,  Cincinnati,  Ohio 
'Meets:  Second  Tuesday  in  Month 
President,  H.  N.  KITCHELL 

4528  Circle  Avenue 
Secretary,  E.  B.  ROYER 
6635  Iris  Avenue 

Illinois 

Headquarters,  Chicago 

Meets:  Second  Monday  in  Month 
President,  R.  E.  HATTIS 

180  N.  Michigan  Avenue 
Secretary,  L,  S.  R*ES 

5614  Blackstone  Avenue 

Kansas  City 

Headquarters,  Kansas  City,  Mo. 

Meets:  Second  Monday  in  Month 
President,  L.  A.  STKPHKNSON 

409  East  13th  Street 
Secretary,  L.  R.  CHASE 

217  Dwight  Building 

Massachusetts 

Headquarters,  Boston 

Meets;  Second  Monday  in  Month 
President,  R.  S,  FRANKLIN 

88  Chauncy  Street 
Secretary.  W.  A.  McPituKsoN 

86  Dwinneli  Street 

West  Roxbury,  Maaa. 

Michigan 

Headquarters,  Detroit 

Meets:  First  Monday  after  the  10th  of  the  Month 
President,  G.  D.  WINANS 

2000  Second  Avenue 
Secretary,  W.  f .  ARNOLD Y 

2847  Grand  Riv<?r  Avenue 

Weatem  Michigan 

Headquarters,  Grand  Rapids 

Meets;  Second  Monday  in  Month 
President,  S.  H,  DOWNS 

211  Creaton  Avenue 

Kalamaasoo,  Mich, 
Secretary*  W,  G.  SCHLICHTINU 

11417  W.  Lovell  Street 

Kalamaasoa,  Mich, 

Minnesota 

Headquarters,  Minneapolis 

Meets;  Second  Monday  in  Month 
President*  C.  E.  LEWIS 

820  Second  Avenue.  S. 
Secretary,  R,  E*  BACKSTEOM 

643  S,  Snelling  Avenue,  St.  Paul,  Minn. 


New  York 

Headquarters,  New  York 
Mt:els:  Third  Monday  in  Month 
President,  H.  W.  FIEDLER 

489  Fifth  Avenue 
Secretary,  T.  W.  REYNOLDS 
100  Pinecrost  Dr., 
Hastings-cm-Hudson,  N.  V. 

Western  New  York 

Headquarters,  Buffalo 
Meets:  Second  Monday  in  Month 

President,  .).  J.  YAGER 

425  Woodbrldfie  Avenue 

Secretary,  P.  S.  HEDLEY 
Curtiss  Building 

Ontario 

Headquarters,  Toronto,  Canada 
Meets;  First  Monday  Every  Other  Month 

President,  W.  R.  BLACKHALL 
3£2  Waverly  Road 

Secretary,  H,  R,  Rcmt 
1104  Bay  Street 

Pacific  Northwest 

Headquarters,  Seattle,  Wash. 

fleets:  Second  Tuesday  in  Month 
President,  A,  L.  POLLARD 

001  Electric  Building 
Secretary,  S.  D,  PBTKRHON 

473  Colman  Building 

Philadelphia 

Headquarters,  Philadelphia,  Pu, 
Meets:  Second  Thursday  in  Month 
President,  W.  P,  CULBKRT 
2019  Rittcnhouee  Street 


Secretary,  W.  R. 

4210  Sansom  Street 

Pittsburgh 

Uu&dituartere,  Pittsburgh,  Pa. 

Meets:  Second  Monday  in  Month 
President,  L,  B.  PITTOCK 

421)  B  Oliver  Building  * 

Secretary,  T.  K.  ROCKWELL 

Carnegie  Institute  of  Technology 

St.  Louis 

Headquarters,  St,  Louis,  Mo, 
Meek;  'First  Wednesday  in  Month 

President,  J,  W»  COOPER 
1590  Arcade  Building 

Secretary,  A,  L,  WALTERS 

7284  Richmond  Place,  Maple  wood,  Mo* 

Southern  California 

Headquarters,  Los  Angeles 
Meets;  Second  Tuesday  in  Month 

President,  W.  K.  BARNUM 
5051  Santa  Fe  Avenue 

Seertiaryt  P.  C.  SCQFIKLD 

748  E,  Washington  Boulevard 

Wisconsin 

Headquartera,  Milwaukee 
Mseh;  Third  Monday  in  Month 

President,  EltKEST  SfcEKRLY 
1817  South  66th  Street 

Secretary,  G.  E,  HOCESTISIN 
3000  W,  Montana  Stettt 


6 


Roll  of  Membership 

AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS 

1934-35 


HONORARY  MEMBERS 

BALDWIN,  WM.  J.  (1915),  New  York,  N.  Y.  (Deceased  May  7,  1924.) 
BILLINGS,  DR,  J.  S.  (1896),  New  York,  N.  Y.  (Deceased  March  10,  1913.) 
GORMLY,  JOHN  (Charter  Member),  Norristown,  Pa.  (Deceased  January  31,  19290 
NEWTON,  C.  W.  (Charter  Member),  Baltimore,  Md.  (Deceased  August  6, 1920.) 
HOOD,  O.  P.  (1929),  Washington,  D.  C. 
JELLETT,  STEWART  A.  (Charter  Member),  (Presidential  Member),  Philadelphia,  Pa. 


LIST  OF  MEMBERS  IN  GOOD  STANDING 
Arranged  Alphabetically— All  Grades 

(Asterisk  indicates  authorship  of  papers) 

(M  1923;  A  1918;  J  1916)  indicates,  Election  as  Member  1923;  Associate  1918;  Junior  1916, 
(Pres.  1923)  indicates,  Elected  President  in  1923  and  is  now  a  Presidential  Member.  * 


ABRAHAM,  Leonard  (J  1035;  S  11)32),  37  S. 

Washington  St.,  Tarry  town,  N.  Y. 
ABRAMS,  Abraham  \M  1927;  J  1924),  Prea,, 

Abbey  Htg,  Co,,  Inc.,  81  Centre  Ave,,  and  (for 

mail),  100  Clove  Rd.,  New  Rochelle,  N.  Y, 
ACHESON,  Albert  R.  (M  19X0),  Consulting  Engn 

(for  mail),  852  Oatrora  Ave,,  Syracuse,  N.  Y. 
ADAMS,  .Benjamin  (M  1919),  Dist.  Mgr.  (for 

mail).  American  Blower  Corp.,  012  Otia  Bldg. 

and  3006  W,  Coulter  St.,  Queen  Lane  Manor, 

Philadelphia,  Pa. 
ADAMS,  Charles  W.  (M  1920),  Salesman,  U,  S. 

Radiator  Corp,,  1405  West  llth  St.,  Kansas 

City,  Mo. 
ADAMS,   Eugene   I.    (M  1934),   Plant   Engr,, 

Michigan  State  College,  and  (for  mail),  115  S. 

Pine,  Lansing,  Mich. 
ADAMS,  Harold  E.  (M  1930),  Chief  Engr.  (for 

mail),  Nash  Engineering  Co.,  South  Net-walk, 

and  Merrill  Heights,  Norwalk,  Conn. 
ADAMS,  Nell  D.  (M  1929;  A  1925:  /  1922),  Supt,, 

Franklin  Htg,  Station  (for  mail),  220  Second 

Ave,  S.W.,  and  836  Eighth  Ave,  &W,,  Rochester, 

Minn, 
ADD  AM  S,  Homer  (Charts?  M&mber;  life  Mm- 

fcr)i  tPrtstotnttol  Umber),  Pres.,  1924;  1st  Vice- 

Pres»,  1928;  Treas,,  1916-1922;  Council,  1915- 

1926) ,  Fre®,,  Kewanee  Boiler  Co.,  Inc.,  and 

FiUgibbont  Boiler  Co,,  Inc-,  570  Seventh  Ave,, 

New  York,  N,  Y, 
ADLAM,  T,  Napier  (M  1982),  Chief  Bnnr.,  Sarco 

Co,,  Inc.,  188  Madison  Ave.,  New  York,  N.  Y,, 

and  (for  mail),  6  Lowell  Ave*,  West  Orange,  N.  J, 
ADtBR,  Alpho*x»e  A,*  (M  1921),  Consulting 

Bnjr.t  33  Stewart  Ave,,  Arlington,  N.  J, 
AEBmY,  John  J,*  (if  1928),  Chief  of  DIv.  of 

Heating,  Ventilation  and  Industrial  Sanitation, 

Chicago  Board  of  Health,  704  City  Hill,  and  (for 

mail),  6821  N.  Oak  Park  Ave.,  Norwood  Park 

F,  (X,  Chicago,  HI. 


AHEARN,  William  J.  (M  1920),  Heating  and 

Ventilating  Engr.,   21   Lake  Rd.,   Cochituate, 

Mass. 
AHLBERG,  Henry  B.  (J  1933),  Engr.,  7  First 

Ave.,  and  (for  mail),  Chase  Brass  &  Copper  Co., 

Waterbury,  Conn. 
AHLFF,  Albert  A.  (M  1923;  A  1918),  Chase  Brass 

&  Copper  Co.,  and  (for  mail),  805  Cook  St, 

Waterbury,  Conn. 
AIKEN,  Jack  F.  (S  1935),  312  Walnut  St.,  S.E., 

Minneapolis,  Minn, 
AKERS,  George  W.  (M  1929),  Secy-Treaa.  (for 

mail),  George  W.  Akers  Co.,  2847  Grand  River 

Ave.,  Detroit,  and  424  WllHtts,  Birmingham, 

Mich. 
ALFSEN,  Nikolai  (M  1933),  Alfsen  &  Gunderson, 

P.  0,  Box  676,  Oslo,  and  (for  mail),  Shabekk  near 

Oslo,  Norway. 
ALGREN,  Axel  B.*  (U 1930),  Inst,  Mech,  Engrg., 

University  of  Minnesota,  Exp,  Engrg.  Lab.,  and 

(for  mail),  6109-17th  Ave.  S.,  Minneapolis,  Minn. 
ALLAN,  Norman  J.  (J  1934),  Asst.  to  Pres., 

Kansas  City  Pump  Co.,  1314  West  llth  St.,  and 

(for  mail),  3661  Madison  Ave.,  Kansas  City,  Mo. 
ALLMAN,  Norman  S,  (S  1934),  8420  Lake  Ave., 

Cleveland,  Ohio. 
ALLSQP,  Rowland  P.  (J  1934),  Mech.  Engr,  (for 

mail),  Mathers  Haldenby,  Archta.,  96  Bloor  St. 

W.,  and  89  Nevill  Park  Blvd.,  Toronto,  Out., 

Canada. 
ALT,  Harold  L.*  (M  1813),  Bid*  Equip.  Engr,, 

Gibbs  &  Hill.  Penn  Station,  New  Y4?rk,  N,  Y., 

and  (for  null),  18-C  Kearay  St.,  Newark,  N,  J. 
AMES,  Charles  F.  (A   1928),  Vice-Pres,  Ames 

Pump  Co.,  Inc.,  30  Church  St,  (for  mail),  Hotel 

Walton,  104  W.  70th  St.,  New  York,  N.  Y* 
AMMERitAN,  Charles  R,  (M  1916),  Consulting 

Engr,  (for  mail),  7724  Century  Bldg.,  and  8908 

GwHord  Ave,,  Indianapolis.  Ind. 
AN0EREGG,  ft.  H,  (M  1920),  Vice-Pres,,  The 
'  Trane  Co.,  and  (for  mail),  324  North  24th, 

taCwae,  wk 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


ANDERSON,  G.  S.  (M  1920),  Mgr.  (for  mail), 
American  Blower  Corp.,  429  Shell  Bldg.,  1008 
W.  Sixth  St.,  and  4267  Holly  Knoll  Dr.,  Los 
Angeles,  Calif. 

ANDERSON,  David  B.  (S  1933),  Engrg.  Dept., 
Wood  Conversion  Co.,  W.  1981  First  National 
Bank,  St.  Paul,  and  (for  mail),  70  Seymour  Ave., 
Minneapolis,  Minn. 

ANDERSON,  Sigurd  H.  (5  1935) ,  3921  Blooming- 
ton  Ave.,  Minneapolis,  Minn. 

ANDES,  William  (A  1934),  Secy-Treas.,  The 
Andico  Co.,  565  Stones  Levee,  and  (for  mail), 
4043  West  103rd  St.,  Cleveland,  Ohio. 

ANDREWS,  George  H.  (A  1934),  Partner,  Frank 
P.  Andrews  &  Son,  and  (for  mail),  213  Meyer 
Ave.,  New  Castle,  Pa. 

ANGUS,  Harry  H.*  (M  191S),  (Council,  1927- 
1929),  Consulting  Engr.,  1221  Bay  St.,  and  (for 
mail),  34  Farnham  Ave.,  Toronto,  Ont.,  Canada. 

ANKER,  George  W.  (J  1935;  S  1933),  Gasoline  & 
Fuel  Oil,  Rensslaer,  and  (for  mail),  27  Jeannelte 
St.,  Albany,  N.  Y. 

ANTHES,  Lawrence  Lee  (A  1935),  Pres.,  Im- 
perial Iron  Corp.,  Ltd.,  30  Jefferson  Ave.4  and 
(for  mail),  64  Jefferson  Ave.,  Toronto,  Ont., 
Canada. 

ARCHDEACON,  Howard  K.  (J  1935;  S  1933),  28 
Niles  PI,,  Yonkers,  N.  Y. 

ARCHER,  David  M.  (M  1934),  Sales  Repr.  (for 
mail),  Sarco  Co.,  Inc.,  143  Federal  St.,  Boston, 
and  11  Wolfe  St.,  W.  Roxbury,  Mass. 

ARENBERG,  Milton  K.  (A  1920),  Dist.  M#r.  (for 
mail),  Ilg  Electric  Vtg.  Co.,  182  N.  LaSalle  St., 
Chicago,  and  1033  S.  Linden  Ave.,  Highland 
1  Park,  111. 

ARMSPAGH,  Otto  W.*  (M  1919),  Chief  Engr., 
Kroeschell  Engrg.  Co.,  2306  N.  Knox  Ave., 
Chicago,  and  (for  mail),  205  S.  Summit  Ave., 
Villa  Park,  111. 

ARMSTRONG,  Robert  W.  (S  1935),  2809  E. 
Lake  of  the  Isles  Blvd.,  Minneapolis,  Minn. 

ARNOLD,  Edward  Y.  (A  1931),  Mgr.  (for  mail), 
Plbg.  &  Htg.  Assns.,  2324  Hampden  Ave.,  and 
1634  Laurel  Ave.,  St.  Paul,  Minn. 

ARNOLD,  Robert  S.  (A  1020;  /  1922),  Dist. 
Mgr.,  Hi  jet  Heater  Sales,  The  Herman  Nelson 
Corp,,  130  South  17th  St.,  Philadelphia,  and  (for 
mail),  Wallingford,  Pa. 

ARNOLD  Y,  William  F.  (A  1930),  Branch  Mgr. 
(for  mail),  2847  Grand  River  Ave.,  Detroit,  and 
520  St.  Uair  Ave.,  Grosse  Points  Village,  Mich. 

ARROWSMITH,  John  O.  (M  1934),  Plant  Engr. 
(for  mail),  Canadian  Kodak  Co,,  Ltd.,  and  389 
Durie  St.,  Toronto,  9,  Ont.,  Canada, 

ARTHUR,  John  M.,  Jr.  (M  1923),  Supt,  Com- 
mercial Light  £  Steam  Sales  (for  mail),  Kansas 
City  Power  &  Light  Co,,  1330  Baltimore, 
Kansas  City,  Mo.,  and  3311  State  Ave.,  Kansas 
City,  Kans. 

ASHLEY,  Carlyle  M.*  (M  1931),  Div.  of  Research 
(for  mail).  Carrier  Bngrg,  Corp.,  750  Fre» 
linghuyaen  Ave.,  Newark,  and  7  GJrard  PI., 
Maplewood,  N,  J. 

ASHLEY,  Edward  E.  (M  1912),  Consulting  Engr, 
(for  mail),  10  East  40th  St.  New  York,  N,  Y., 
and  P.  O.  Box  188,  Noroton  Heights,  Conn. 

ASTON,  James  (M  1019),  A,  M.  Byers  Co.,  235 
Water  St.,  Pittsburgh,  Pa. 

ATHERTON,  G.  R.  (M  1030),  40  West  40th  St., 
New  York,  N,  Y. 

ATKINS,  Thomas  3.  (M  1931),  Sales  Engr., 
Carrier  Engrg.  Corp.,  12  South  12th  St,  Phila- 
delphia, and  (for  mail),  119  Kenllworth  Rd., 
Merlon,  Pa, 

ATKINSON,  Kenneth  B.  (J  1930),  &  Eppirt  St, 
East  Orange,  N.  J, 

AVEKY,  tester  T  (M  1934),  Free,  (for  mail), 
Avery  Engrg.  Co.,  2341  CarnegieAve.,  Cleveland, 
and  21149  Colby  Rd,,  Shaker  Heights,  Ohio. 

AXEMAN,  James  E.  (M  1032;  A  1081;  J  1925), 
Br,  Mgr*  (for  mail),  Spencer  Heater  Co.,  1205 
Court  Square  Bldg.,  and  908  Old  Oak  Rd,, 
Stoneleigh,  Baltimore,  Md. 


B 

BACHLER,  Leonard  J.  (M  1918),  304  East  41st 

St.,  New  York,  N.  Y. 
BACKSTROM,  Russell  E.. (A  1931;  J  1928),  (for 

mail),  Wood  Conversion  Co.,  1st  National  Bank 

Bldg.,  and  543  S.  Snelling  Ave.,  St.  Paul,  Minn. 
BACKUS,  Theodore  H.  L.  (M  1916),  Schumacher 

&  Backus,  200-208  Hill  St.,  Ann  Arbor,  Mich. 
BADGETT,    W.    Howard*    (J    1932),    Research 

Asst.,  Texas  Engrg.  Experiment  Station,  College 

Station,  Texas. 
BAHNSON,  Frederick  F.*  (M  1917),  Vice- Pres. 

and  Chief  Engr.   (for  mail),  The  Bahnson  Co., 

1001   S.    Marshall   St.,   and   28   Cascade  Ave., 

Winston  Salem,  N.  C. 
BAILEY,  Edward  P.,  Jr.  (M  1925),  Consultant, 

Mayfield    Rd.   at   Lee    Blvd.,   and    (for   mail), 

2475  Lee  Blvd.,  Cleveland,  Ohio. 
BAILEY,    W.    Mumford    (M    1930),    Managing 

Director,  Mumford  Bailey  &  Preston,  Ltd.,  and 

Joint    Managing   Director,    British   Trane    Co,, 

Ltd.  (for  mail),  "Newcastle  House,"  Clcrkenwell 

Close,    London    EC1,    and    "Oldbury    Court," 

Dainesway,  Thorpe  Bay,  Essex,  England. 
BAKER,  Howard  C*  (M  1921),  The  II.  C.  Baker 

Co.,  12S  S.  St.  Clair  St.,  Toledo,  Ohio. 
BAKER,  Irving  G.   (KI  1921),  Mgr.  Air  Cond. 

Div,  (for  mail),  York  Ice  Machinery  Corp.,  and 

004  Linden  Ave.,  York,  Pa. 
BAKER,  Roland  H.  (M  1928;  A  1924),  Pres.  (for 

mail),  R,  H.  Baker  Co.,  Inc.,  145  Broadway,  and 

420  Memorial  Dr.,  Cambridge,  Muss. 
BALDWIN,    William   Howard    (M    1921),    Br. 

Mgr.  (for  mail),  C.  A.  Dunham  Co.,  2988  1C. 

Grand  Blvd.,  and  1022  Virginia  Park,  Detroit, 

Mich. 
BALSAM,  Charles  P.  (M  1932),  324  Fourth  St., 

Brooklyn,  N.  Y. 
BARBERA,  Henry  A.  (S  1932),  1727  Colden  Ave., 

New  York,  N.  Y. 
BARBIERI,  Patrick  J.  (S  1933),  2100  Belmont 

Ave.,  New  York,  N.  Y. 
BARNES,  Walter  E.  (M  1933),  Pres.,  Barnes  & 

Jones,  Inc.,  128  Brookaide  Ave.,  Jamaica  Plain, 

Boston,    and    (for   mail),    7   Woodlawn   Ave.. 

Wellesley  Hills,  Mass. 
BARNS,  Amos  A,  (M  1933),  Owner  (for  matt), 

440  W.  State  St.,  Ithaca,  N.  Y. 
BARNUM,  Charles  R,   (3  1935),  1494  Capitol 

Ave.,  St.  Paul,  Minn.  , 

BARNUM,  Marvin  C.  (M  1930,-  A  1928),  Rm. 

1022-1133  Broadway,  New  York,  N.  Y. 
BARNUM,  Willis  E.,  Jr.  (M  1933;  A  1933;  J 1930) 

Sales  Engr.,  York  Ice  Machinery  Co.,  5061  Santa 

Fe  Ave,,   Los  Angeles,   and  (for  mail),   249(1 

Poplar  PL,  Huntington  Park,  Calif. 
BARR,  Geor&e  W.  (M  1905).  0!at.  M«r,,  Aeroan 

Corp.,  Land  Title  Bldg,,  Philadelphia,  and  (for 

mail),  Woods  End,  Villanova,  Pa. 
BARRY,  James  G.,  Jr.  (M  1933),  VJce-Pres.  (for 

mail),  'Elliott  &  Barry  Engr«,  Co.,  4060  W.  Pine 

Blvd.,  and  5051  Queens  Ave,,  St.  Louis,  Mo. 
BARRY,  Patrick  L  (M  1920),  M.  Barry,  Ltd.,  4 

Marlboro  St.,  Cork,  Ireland. 
BARTH,    Herbert   fe.    (M    1920L    Sales    Mgr., 

American    Blower    Corp.,    6000    Russell    St., 

Detroit,  Mich, 
BARTLETT,  Amos  C.  (M  1919),  Diat,  Mgr.  (for 

mail),   B,   F.  Sturtevant  Co,,   89   Broad   St., 

Boston,  and  30  HoIlingBworth  Ave.,  Braintree, 

Mass. 
BARTLETT,  C.  Edwin  (M  1922),  Pnss.,  Bartlett 

&  Co.,  Inc.  (for  malt),  1938  Market  St.,  and  Sill 

W.  Coulter  St.,  Philadelphia,  Fa. 
BASTEDO,  Albert  E,  (M  1919),  Vice-Pres-Treaa- 

Mgr.  (for  mall),  Burnham  Boiler  Corp.*  Irving- 

ton-on-Hudson,  and  Burnslde  Dr.,  Hastings-on- 

Hudson,  N.  Y. 
BAUM,  Albert  L.  (M  1916),  Member  of  Firm  (for 

mail),  Jaros  Baum  &  Ballet,  415  Lexington  Ave»» 

and  001  West  113th  St.,  New  York,  N.  Y» 
BAUMGAR0NEH,  Carroll  Mile*  (M  1028),  $r. 

Mar.  (for  mall),  U.  S.  Radiator  Corp.,  S254  N, 

Kirbourn  Ave.»  Chicago,  and  602  Michigan  Ave., 

Evanston,  111. 


ROLL  OF  MEMBERSHIP 


BAYSE,  Harry  V.  (M  1923),  American  Furnace 
Co.,  2725  Morgan  St.,  St.  Louis,  Mo. 

BEARD,  Earl  L.  (S  1934),  736  East  13th  St., 
Oklahoma  City,  Okla. 

BEAURRIENNE,  Auguste*  (M  1912),  Consulting 
Engr.,  25  Rue  des  Marguettes,  Paris,  France. 

BEAVERS,  George  R.  (M  1929),  Chief  Engr., 
Canadian  Blower  &  Forge  Co.,  Ltd.,  Woodside 
Ave.,  and  (for  mail),  168  Samuel  St.,  Kitchener, 
Ont.,  Canada. 

BEEBE,  Frederick  E.  W.  (A  1915),  Johnson 
Service  Co.,  28  East  29th  St.,  New  York,  N.  Y. 

BEGGS,  William  E.  (M  1927),  Pres.,  W.  E.  Beggs 
Co.,  907  Lloyd  Bldg.,  and  (for  mail),  3639 
Palatine  Ave.,  Seattle,  Wash. 

BEIGHEL,  Howard  Atlee  (A  1927),  Sales  Repr. 
(for  mail),  The  Herman  Nelson  Corp.,  503 
Columbia  Bank  Bldg.,  Pittsburgh,  and  207 
Puritan  Rd.,  Rosslyn  Farms,  Carnegie,  Pa. 

BEITZELL,  Albert  E.  (A  1933;  J  1930),  Mgr., 
Westinghouse  Air  Cond.  Div.  of  Wm.  E.  Kings- 
well,  Inc.,  1214-24th  St.,  and  (for  mail),  1339 
Girard  St.  N.W.,  Washington,  D.C. 

BELING,  Earl  H.  (A  1930;  J  1925),  2428-13th 
St.,  Moline,  111. 

BELL,  E.  Floyd  (M  1933),  (for  mail),  619  Foshay 
Tower,  and  2605  Fremont  Ave.  S.,  Minneapolis, 
Minn. 

BEMAN,  Myron  C.  (M  1926),  (Council,  1934), 
Consulting  Engr.  (for  mail),  Beman  &  Candee, 
374  Delaware  Ave.,  and  699  Richmond  Ave., 
Buffalo,  N.  Y. 

BENNETT,  Edwin  A.  (J  1929),  Sales  Engr.  (for 
mail),  American  Blower  Corp.,  401  Broadway, 
New  York,  and  45  Pondfield  Rd.  W.,  Bronxville, 
N.  Y. 

BENNITT,  George  E.  (M  1918),  Consolidated 
Gas  Co.  of  New  York,  4  Irving  PI.,  New  York, 
N.  Y. 

BENOIST,  LeRoy  L.  (M  1934),  Mgr.  (for  mail), 
Benoist  Bros.  Hardware  &  Sup.,  117  South  10th 
St.,  and  1500  Main  St.,  Mt.  Vernon,  111. 

SENSE,  William  M.  (S  1934),  Engr.,  Institute 
of  Tlaermal  Research  (for  mail),  American 
Radiator  Co,,  675  Bronx  River  Rd.,  Yonkers, 
and  340  Hayward  Ave.,  Mt.  Vernon,  N.  Y. 

BENTZ>  Harry  (M  1915),  18  Holland  Terrace, 
Montclair,  N.  J. 

BERCHTOLD,  Edward  W.  (M  1927;  A  1925) , 
Rate  Engr,  (for  mail),  Boston  Consolidated  Gaa 
Co.,  100  Arlington  St.,  Boston,  and  20  Randolph 
St.,  S.  Weymouth,  Mass. 

BERGHOEFER,  Victor  A.  (/  1926),  Vice-Pres., 
Sterling  Engrg.  Co.,  3738  N.  Holton,  and  (for 
mail),  4129  North  20th  St,,  Milwaukee,  Wis. 

BERMAN,  Louis  K.  (M  1008),  Pres.  (for  mail), 
Raisler  Heating  &  Sprinker  Cos.,  129  Amsterdam 
Ave.,  and  101  Central  Park  West,  New  York, 
N.  Y. 

BERMEL,  Alfred  H.  (A  1933;  J  1928),  16  Pershing 
PL,  North  Arlington,  N.  J. 

BERNHAR0,  Georfte  (A  1929),  Pres.,  Bernhard 
Engrg.  Corp,,  101  Park  Ave.,  New  York,  and 
(for  mail),  18  Liamore  Rd.,  Lawrence,  L,  L,  N.  Y. 

BERNSTROM,  Bert  (M  1930),  Engr.,  132  West 
04th  St.,  New  York,  N.  Y, 

BEST,  Milliard  W,  (A  1933) f  Pres,  (for  mail), 
KoMeetric  Underfeed  Stoker  Co.,  Ltd,,  245 
Kenilworth  Ave.  S,,  and  1750  King  St.  K., 
Hamilton,  Ont,  Canada. 

BETLBM,  Henrietta  T,  (J  1984),  (for  mall), 
Betlem  Heating  Co.,  1026  Eaat  Ave.,  and  1293 
Park  Ave.,  Rochester,  N.  Y, 

BETTS,  Howard  M,  (U  1927),  Senior  Mech, 
Kngr.,  Htg,  &  Vtg.  (for  mail),  Dept.  of  Bldp,, 
City  of  Minneapolis,  213  City  Hall,  and  4923 
Russell  Ave.  S,,  Minneapolis,  Minn. 

BETZ,  Harry  D.  (M  1928),  Pres.  (for  mail),  Betz 
Unit  Air  Cooler  Co.,  6  W.  Ninth  St.,  and  4210 
Mercer,  Kansas  City,  Mo, 

BILYEU,  William  F,  (M  1927),  Eastern  Div. 
Mgr,  (for  mail),  The  Trane  Co,,  1109  Chanln 
Bldg.,  New  York,  and  Gibson  Apt.,  Flushing, 
L*  I*  N.  Y.. 


BINDER,  Charles  G.  (M  1920),  Mgr.  Htg.  Dept., 
Warren  Webster  &  Co.,  17th  and  Federal  Sts., 
Camden,  and  (for  mail),  115  Oak  Terrace, 
Merchantville,  N.  J. 

BINFORD,  Wilmer  M.  (J  1930),  Mgr.  Contract 
Dept.,  S.  Div.  (for  mail),  2120  East  25th  St.,  and 
6215  San  Vicente  Blvd.,  Los  Angeles,  Calif. 

BIRD,  Charles  (A  1934),  Treas.  and  Gen.  Mgr. 
(for  mail),  The  Doermann-Roehrer  Co.,  450-56 
E.  Pearl  St.,  and  3026  Beaver  Ave.,  Cincinnati, 
Ohio. 

BIRRELL,  Allan  L.  (A  1925),  Consulting  Engr., 
372  Bay  St.,  Toronto  2,  and  (for  mail),  93 
Kingsway,  Toronto  9,  Canada. 

BISCH,  Bernard  J.  (M  1931),  Engr.,  St.  Mary  of 
The  Woods  College,  St.  Mary  of  The  Woods,  Ind. 

BISHOP,  Charles  R.  (Life  Member;  M  1901)  ,413 
Locust  St.,  Lockport,  N.  Y. 

BISHOP,  Frederick  R.  (M  1921),  8011  Dexter 
Blvd.,  Detroit,  Mich. 

BJERKEN,  Maurice  H.  (A  1927),  Dist.  Repr.  (for 
mail),  Hoffman  Specialty  Co.,  533  S.  Seventh  St., 
and  4952-17th  Ave.  S.,  Minneapolis,  Minn. 

BLACK,  Edgar  N.,  3rd  (M  1922),  Philadelphia 
Mgr.,  Fitzgibbons  Boiler  Co.,  Inc.,  814  Land 
Title  Bldg.,  Philadelphia,  and  (for  mail).  111 
Woodside  Rd.,  Haverford,  Montgomery  Co.,  Pa. 

BLACK,  F.  C.  (M  1919),  Pres.  (for  mail),  F.  C. 
Black  Co.,  622  W,  Randolph  St.,  and  4535  N. 
Ashland  Ave.,  Chicago,  111. 

BLACK,  Harry  G.  (M  1917),  Prop,  (for  mail), 
P.  Gormly  Co.,  155  North  10th  St.,  and  927 
North  65th  St.  Philadelphia,  Pa. 

BLACK,  William  B.  (J  1932),  Bryant  Heater  Co., 
135  Seward  Ave.,  Bradford,  Pa. 

BLACKBURN,  Edwin  C.,  Jr.  (M  1929),  Con- 
sulting Engr.,  12  Clermont  Ave.,  Hempstead, 
L.  L,  N.  Y. 

BLACKBALL,  Wilmot  R.  (M  1922),  Partner, 
McKellar  &  Blackball,  1104  Bay  St.,  and  (for 
mail),  332  Waverly  Rd.,  Toronto,  Canada. 

BLACKMAN,  Alfred  O.  (M  1911),  Consulting 
Engr.  (for  mail),  145  West  45th  St.,  and  149 
West  12th  St.,  New  York,  N.  Y. 

BLACKMORE,  F.  H.  (M  1923),  Mgr.  Operating 
Dept.  (for  mail),  U.  S.  Radiator  Corp.,  Box  686, 
Detroit,  and  515  Tooting  Lane,  Birmingham, 
Mich. 

BLACKMORE,  Georfce  C.  (Charter  Member; 
Life  Member},  Pres.,  Automatic  Gas  Steam 
Radiator  Co.,  301  Brushton  Ave.,  Pittsburgh,  Pa. 

BLACKMORE,  J.  J.*  (Charter  Member:  Life 
Member),  32  West  40th  St.,  New  York,  N,  Y/ 

BLACKMORE,  James  S,  (J  1931),  Sales  Engr., 
H.  A.  Thrush  &  Co.,  Peru,  Ind.,  and  (for  mail), 
4315  Maple  Ave,,  Edgewood,  Pittsburgh,  Pa. 

BLACKSHAW,  J.  L,*  (J  1929),  68  Plaza  St., 
Brooklyn,  N.  Y. 

BLANDmG,  George  H.  (M  1919),  800  N. 
Lombard  Ave.,  Oak  Park,  111. 

BLANKIN,  Merrill  F.  (M  1927;  A  1926;  J  1919) 
Pres.  (for  mail),  Haynes  Selling  Co.,  Inc.,  1518 
Fairmount  Ave,,  and  3328  W.  Penn  St.,  Phila- 
delphia, Pa. 

BLISS,  Georft©  L.  (A  1933),  Engr.  and  Sales,  (for 
mail),  Allia-Chalmers  Mfg.  Co.,  1410  Waldheim 
Bldg.,  llth  and  Main,  and  7041  Brooklyn  Ave., 
Kansas  City,  Mo. 

BOALES,  William  G.  (A  1923),  Mfr.  Agt  (for 
mail),  6537  Hamilton  Ave.,  Detroit,  and  195 
McMillan  Rd.,  Grosse  Points  Farms,  Mich, 

BOCK,  Bernard  A.  (A  1929;  /  1927),  Engrg. 
Draftsman,  425  Beech  St.,  Arlington,  N,  J. 

BOCK,  L  I.  (A  1934),  Sales  Engr.  (for  mail),- 
Carrier  Engrg.  Corp,,  2022  Bryan  St.,  and  2500 
South  Blvd.,  Dallas,  Texas. 

BODDINGTON,  William  P.  (M  1927),  Mgr.  (for 
mail),  The  Canadian  Powers  Regulator  Co., 
Ltd.,  106  Lombard  St.,  and  280  Clendenan  Ave.» 
Toronto,  Ont,,  Canada. 

BODINGBR,  J.  H.  (M  1931),  Prea.  (for  mail), 
Bodinger  &  Co,,  Inc.,  439  West  38th  St,»  New 
York,  and  1429  East  19th  St.,  Brooklyn,  N.  Y, 

BOGATY,  Hermann  S.  (M  1921),  5230  North 
l»th  St.,  Philadelphia,  Pa, 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


BOLSINGER,  Raymon  C.  (M  1916),  Mgr.  (for 
mail),  Automatic  Florzone  Htg.  Co.,  319  E.  Main 
St.,  Norristown,  Pa.,  and  238  E.  Madison  Ave., 
Collingswood,  N.  J. 

BOLTE,  E,  Endicott  (A  1929),  Salesman, 
National  Radiator  Corp.,  1111  East  83rd  St.,  and 
(for  mail),  6516  Kenwood  Ave.,  Chicago,  111. 

BOLTON,  Reginald  Pelham*  (Life  Member; 
M  1897),  (Presidential  Member},  (Pros.,  1911; 
1st  Vice-Pres,  1905-1910;  2nd  Vice-Pres.,  1903; 
Board  of  Governors,  1901,  1905,  1910,  1911, 
1912,  1913),  The  R.  P.  Bolton  Co.,  116  East  19th 
St.,  New  York,  N.  Y. 

BOLZ,  Harold  A.  (J  1934),  Asst.  Instructor, 
Mech.  Bngrg.  Dept.,  Case  School  of  Applied 
Science,  University  Circle,  Cleveland,  and  (for 
mail),  3558  East  159th  St.,  Shaker  Heights,  Ohio. 

BOND,  Horace  A.  (M  1930),  Mgr.,  Warren 
Webster  Co.,  91  State  St.,  and  (for  mail),  12 
Ramsey  PI.,  Albany,  N.  Y. 

BOOTH,  C.  A.  (M  1917),  Vice-Pres.  (for  mail), 
Buffalo  Forge  Co.,  490  Broadway,  and  142 
Summit  Ave.,  Buffalo,  N.  Y. 

BOOTH,  Harry  N.  (M  1924;  A  1917),  Vice-Pres. 
Sales  Dept.  (for  mail),  U.  S.  Radiator  Corp., 
Room  1056  1st  National  Bank  Bldg.,  and  688 
Taylor  Ave.,  Detroit,  Mich. 

BORLING,  John  R.  (A  1934),  Engr.-Custodian 
Board  of  Education,  6520  S.  Wood  St.,  and  (for 
mail),  6818  Normal  Blvd.,  Chicago,  111. 

BORNEMANN,  Walter  A.  (M  1924;  J  1923), 
Sales  Engr.  (for  mail),  Carrier  En«r.  Corp.,  12 
South  12th  St.,  Philadelphia,  and  123  W. 
Wharton  Ave.,  Glenside,  Pa. 

BORUCH,  Edwin  R.  (A  1935),  Sales  Engr,  (for 
mail),  Dallas  Power  &  Light  Co.,  1506  Com- 
merce, and  835  N.  Bishop,  Dallas,  Texas. 

BOUCHERLE,  Henry  N.  (M  1934),  Secy,  (for 
mail),  The  Scholl-Chofrln  Co.,  Mahoning  Ave. 
and  Ho&ue  St.,  and  3412  Hudson  Ave.,  Youngs- 
town,  Ohio. 

BOUEY,  An&us  J.  (J  1930),  Sales  Engr,  (for  mail), 
B.  F.  Sturtevant  Co.,  553  Monadnock  Bldg., 
and  4810  Fulton  St.,  San  Francisco,  Calif. 

BOUILLON,  Lincoln  (M  1933),  Consulting  Engr., 
1411  Fourth  Ave,  Bldg,,  and  (for  mail),  4186- 
42nd  Ave.  N.E.,  Seattle  Wash. 

BOWERS,  Arthur  F.  (A  1910),  Pres.,  Industrial 
Htg.  &  Engrg,  Co.,  828  N.  Broadway,  Milwaukee, 
Wis. 

BOWERS,  Ross  C.  (A  1932),  Br.  Mgr.  (for  mail), 
Minneapolis-Honeywell  Regulator  Co.,  335  W. 
North  Ave.,  and  3773  North  52ncl  St.,  Mil- 
waukee, Wis, 

BOWLES,  Potter  (A  1928),  Pres.  (for  mail), 
Hoffman  Specialty  Co.,  Inc.,  500  Fifth  Ave., 
Room  3324,  New  York,  and  078  Ely  Ave., 
Pelham  Manor,  N.  Y. 

BOWMAN,  James  W.  (S  1934),  210  S,  Saute  He, 
Norman,  Okla. 

BOYDEN,  Davis  S.*  (M  1900),  (Council,  1930-34; 
Treaa.,  1983-34),  Stint.,  Steam  Htg,  Service 
Dept,  (for  mail),  Etlison  Electric  Illuminating 
Co.  of  Boston,  30  Boylaton  St.,  Boston,  and  1406 
Commonwealth  Ave,,  Brighton,  Mass. 

BOYKER,  Robert  Owen  (J  1935),  Partner  (for 
mall),  Mac  Boyker  &  Son,  and  102  Kennebeck 
Ave.»  Kent,  Wash, 

BRAATZ,  Chester  Johnson*  (M  1930),  Ensrn 
Temp.  Control,  Barber-Colman  Co,,  and  (for 
mail),  718  King  St.,  Rockford,  111, 

BRABBLE,  Dr.  Charles  W.*  (M  1925),  Dir.t 
Institute  of  Thermal  Research  (for  mail), 
American  Radiator  Co.,  07U  Bronx  River  Rd., 
Yonkera,  and  Weatchester  Park,  50  Lincoln 
Ave.,  Tuckahoe,  N,  Y. 

BRACKEN,  John  Henry  (M  1927) »  Mgr., 
Industrial  Uses  Dept.  (for  mail),  The  Cdotftx 
Co.,  919  N,  Michigan  Ave,,  Chicago,  III 

BRADFIELD,  William  W,  (M  10263,  Consulting 
Engr.  (for  mail),  i)01  Michigan  Trust  Bldg,,  and 
18&  Franklin  St,  S,B,,  Grand  Rapids,  Mich. 

BRADLEY,  Eugene  I*.  (M  1900),  Fres.  (for  mail), 
Hester-Bradley  Co.,  2835  Washington  Ave., 
and  G98#  Penning  Ave.,  St.  Louis,  Mo. 


BRAEMER,  William  G.  R.  (M  1915),  (for  mail), 
Wm.  G.  R.  Braemer  &  Josiah  H,  Smith,  Engrs, 
Room  1265  Commercial  Trust  Bldg.,  Phila- 
delphia, Pa.,  and  223  Chestnut  St.,  Haddon- 
field,  N.  J. 

BRANDI,  O.  H.  (M  1930),  Lufttechnische  Gesell- 
schaft  m.  b.  H.,  Berlin  W.  50,  Nurnbergerstr. 
53/55,  and  (for  mail),  Landoltwcg  21,  Berlin, 
Dahlcm,  Germany, 

BRANDT,  Ernst  H.,  Jr.  (M  1928),  Pres.,  Reliance 
Engrg.  Co.,  Inc.,  515  N.  Church  St.,  and  (for 
mail),  P.  O.  Box  1292,  Charlotte,  N.  C, 

BRAUER,  Roy  (M  1920),  Prop,  (for  mail), 
Ventilating  Equip.  Co.,  1101  Bessemer  Bldg., 
Pittsburgh,  and  R.  F.  D.  No.  1,  Hillcrest,  Library, 
Pa. 

BRAUN,  John  J.  (M  1932),  Factory  Mgr.,  The 
U.  S.  Playing  Card  Co.,  Norwood  Station,  Cincin- 
nati, and  (for  mail),  4305  Floral  Ave.,  Norwood, 
Ohio. 

BRAUN,  Louis  T.  (M  1021),  Executive  Secy,  (for 
mail),  Chicago  Master  Stcamfittcra  Assn.,  228 
N.  LaSalle  St.,  and  1548  Pratt  Blvd.,  Chicago, 

BRECKENRIDGE,  L.  P.*  (Life  Member,-  M  1920),  ' 

The  Brackens,  N.  Ferrisburg,  Vt. 
BRE0ESEN,  Bernhard  P.  (A  1931),  3119  Knox 

Ave.  N.,  Minneapolis,  Minn. 
BREITENBACH,  George  C.  (A/  1933;  A  1933; 

J   1928),    Sales   Engr.,    The   Trane   Co.,   200G 

Chestnut  St.,  Philadelphia,  and  300  Essex  Ave., 

Apt.  203  A,  Narberth,  Pa. 
BRENEMAN,  Robert  B.  (A  1931;  J  1927),  Sales 

Engr.  (for  mail),  Armstrong  Cork  &  Insulation 

Co.,  232  W.  Seventh  St.,  and  1557  Addingham 

PI.,  Cincinnati,  Ohio. 
BRENNAN,  John  W.  (M  1935;  A  1934),  Salesman 

(for  mail),  American   Blower  Corp.,   Ilofmarm 

Bids.,  and  594-1-  Yorkshire,  Detroit,  Mich, 
BRIDE,  William  T.  (M  1928;  A  1928:  J  1025), 

Supt.,  Enurg.  (for  mail),  P.  O.  Box  777,  Lawrence, 

and  50  High  St.,  Mcthucn,  Mass. 
BRIGHAM,  Frederick  H.  (M  1930)  ,  Sales  En«rM 

G.   H.   Gleason  &   Co,,   25   Huntington   Ave,, 

Boston,  and  (for  mail),  80  Bedford  St.,  Lexing- 

ton, Mass, 
BRINKER,  Harry  A.  (Af  193-1),  Member  of  Firm, 

Wilson-  Brinker  Co.,  412  Pythian  Bldg.,  and  (for 

mail),  524  Village  St.,  Kalamausoo,  Mich. 
BRINTON,  Joseph  Ward  (M  1920),  Dist,  Mgr. 

(for  ranil),  American  Blower  Corp.,  1003  Statler 

Bldg,,  Boston,  and  42  Gleason  St.,  West  Mcdford, 

Mass, 
BRISSBTTB,  Leo  A.  (M  1030),  Treaa.  (for  mail), 

Trask  Ht«.  Co,,  4  Merriniac  St.,  Boston,  and  108 

Florence  St.,  Melrose,  Mass. 
BRODERICK,   Edwin   L.*  (M   1033),    Research 

Aast,  in  Mech,  Kn«r,  (for  mail),  University  of 

Illinois,  210  M,  E,  Lab,,  and  1108  W.  Stoughton 

St.,  Urbana,  111. 
BRONSON,  Carlos  E.*  (M  1919),  M«ch,  Engr. 

(for   mail),    Kewan«e    Boiler   Carp,,    and    811 

McKinley  Ave.,  Kewanee,  Ul. 
BROO&S,  Frank  W.  (S  1931),  (for  mail),  2111 

Abington  Rd.,  Cleveland,  and  93S  N,  Broadway, 

Dayton,  Ohio, 
BROOM,  Benjamin  A.  (M  1914),  Sales  Promo- 

tion Engr.,  Weil  McLain  Co.,  Ml  W.  Lake  St., 

and  (for  mail),  1644  Sherwln  Ave.,  Chicago,  111, 
BROWN,  Alfred  F,   (M  1927),   Vice-Fres,    (for 

mail),    Reynolds    Corp,,    809    N,    LaSaiJe    St., 

Chicago,  and  551  Hill  Terrace,  WinnHfca,  111, 
BROWN,  Aubrey  L*  (M  1028),  Prof,  of  Htg,  and 

Vt«,  (for  mail),  Ohio  State  University,  and  I8» 

Richards  Rd,,  Columbus,  Ohio, 
BROWN,    Poufcott*   (M    1986),    VicenFres.    tfor 

mail),  Gray  &  Dudley  Co.,  W  Third  Am  N., 


,  ., 

P,  O,  Bosc  7232,  and  2314  West  Ead 
ville,  Tenn, 


e.,  Naeh* 


BROWN,  Morrlft  (J  1028),  Htg.  Eagr.  (for  mail), 
Brown  Bros.,  340  Talbot  Ave.,  and  00©  Park  St., 
Dorchester,  Maaa, 

BROWN,  Ronald  F.  (S  1033),  ©6  Mitchell  Av«.» 

Blnghamton,  N.  Y, 


10 


ROLL  OF  MEMBERSHIP 


BROWN,  Tom  (M  1930),  Gen.   Mgr.   (for  mail), 

Autovent  Fan  &  Blower  Co.,  1805  N.  Kostner 

Ave.,  and  5826  Lake  St.,  Chicago,  111. 
BROWN,  William  A.   (M  1930),  2523-14th  St. 

N.W.,  Washington,  D.  C 
BROWN,  William  H.  (A  1923),  Mgr.  (for  mail), 

Brown   Bros.,  3310  W.   North  Ave.,  and  3015 

North  22nd  St.,  Milwaukee,  Wis. 
BROWN,  W.  Maynard  (A  1930),  Warren  Webster 

&  Co.,  17th  and  Federal  Sts.,  Camden,  N.  J. 
BROWN,  W,  Murray  (J  1935;  5  1930),  Draftsman 

and  Estimator  (for  mail),  William  P.  Brown,  31 

Sanford  St.,  and  7S  Randolph  St.,  Springfield, 

Mass. 
BROWNE,  Alfred  L.   (M  1923),  Illinois  Engrg. 

Co.,  3514  Grand  Central  Terminal,  New  York, 

N.   Y.,  and  253  Highland  Rd.,  South  Orange, 

N.  J. 
BRUGKMANN,  John  C.  (J  1935;  5  1932),  Sales 

Repr.,  American  Radiator  Co.,  40  West  40th  St., 

and  (for  mail),  2290  Sedgwick  Ave.,  New  York, 

N.  Y. 
BRUEGGEMAN,  Arthur  R.  (M  1920),  (for  mail), 

The  Erie  Engineering  Co.,  1740  East  12th  St., 

Cleveland,    and    17220   Aldersyde    Dr.,    Shaker 

Heights,  Ohio. 
BRUNETT,  Adrian  L.  (M  1923),  Assoc.  Mech. 

Engr,,     U.    S.    Supervising    Architects    Office, 

Treasury  Dept.,  Washington,  D.  .C.,  and   (for 

mail),  P.  O.  Box  36,* Rockville,  Md. 
BRUST,  Otto  (M  19,30) ,  (for  mail) ,  Luf ttechnische 

Gesellschaft,  Prag  1,  Revolucni  13  and  Veverkova 

ul  3  Prag  VII  Czechoslovakia. 
BRYANT,  Dr.  Alice  G.  (M  1921),  502  Beacon  St., 

Boston,  Mass. 
BRYANT,  Percy  J.  (M  1915),  Chief  Engr.   (for 

mail),  Prudential  Insurance  Co.,  783  Broad  St., 

Newark,  and  754  Belvidere  Ave.,  Westfield,  N.  J, 
BUCK,  Lucien  (M  1928),  Pres.  (for  mail),  Buck 

Dryer  Corp.,  P.  O.  Box  308,  Manchester,  Conn. 
BUCKLEY,  Martin  B.  (A  1930),  824  Grand  Ave., 

Kansas  City  Mo. 
BUENGER,  Albert*  (M  1920;  J  1917),  (Council, 

1934),  Mech.  Engr.  (for  mail),  C.  H.  Johnston 

Archt.,    715    Empire    Bank    Bldg.,    and    1606 

Stanford  Ave.,  St.  Paul,  Minn, 
BUENSOD,    Alfred    Charles    (M    1918),    Sales 

Kngr.,  Carrier  Engrg.  Corp.,  Chrysler  Bldg.,  and 

(for  mail),  1  Fifth  Ave.,  New  York,  N,  Y. 
BUFORD,  Jack  W.  (J  1935;  5  1933),  2323  Ash- 
land Ave.,  Walnut  Hills,  Cincinnati,  Ohio. 
BULKELEY,  Claude  A>  (M  1923),  Chief  Engr. 

(for  mail),  Niagara  Blower  Co.,  6  East  45th  St., 

and  410  West  68th  St.,  New  York,  N.  Y. 
BULL,  Alvah  Stanley  (J  1935;  S  1933),  304  West 

35th  St.,  Minneapolis,  Minn. 
BULLEIT,  Charles  R.  (M  1932;  A  1932;  J  1930), 

281-2  Austin  Ave.,  Bvansville,  Ind. 
BULLOCK;,  Howard  H.  (A  1933),  Commercial 

Engr,   (for  mail).  General  Electric  Co.,  5201 

Santa  Fe  Ave,,  Los  Angeles,  and  2530  Grand  St., 

Walnut  Park,  Calif. 
BULLOCK,  Thomas  A,   (M  1980),   Engr.   (for 

mail)*  Densmorc,  LeClear  &  Robbina,  31  St. 

James   Ave,t    Boston,   and   89   Fairmont   St., 

Arlington »  Mass. 
BUOT,  Antonio  V,  (S  1036),  2730  Portland  Ave, 

S,,  Minneapolis,  Minn, 
BUR,  Julian  R,  C.  (J  1931),  Chief  Engr.  (for  mail). 

Bur  &  Co,,  10  Rue  du  Chapeau  Rouge,  and  1 

Place  Francois,  Rude  Dijon,  France.  • 
BURBAUM,  W.  Allen  (J  1933),  Asst.  Br.  Mar., 
Rex  Cok  Inc.,    2392  Grand  Concourse,  New 
York,  and  (for  mail),  180  Clinton  Ave*,  Brooklyn, 
N,  Y. 

BURCH,  Laurence  A.  (M  1934).  Mgr,  Htg,  Div., 
Porfex  Radiator  Corp,,  415  W,  Oklahoma  PL, 

and  (for  mall),  421  B.  uoyd  St,  Milwaukee,  WJa. 
BURKB,  James  J*  (/  1930),  Engr.  (Air  ConcU), 
Carrier  Bngri,  Corp,,  850  Frellnflhuysen  Ave., 
Newark*  and  (for  mall),  720  N.  Broad  St., 
£li*»b«th,  N*  J. 

BURKft,  William  J.  (A  1934),  1109  S.  Cartoon 
8L»  Tulsa,  Old*. 


BURNETT,   Earle   S.    (M   1920),   Mech.   Engr., 

U.  S.  Bureau  of  Mines,  Arnarillo  Helium  Plant, 

P.  O.  Box  2025,  and  (for  mail),  4223  West  llth 

Ave.,  Amarillo,  Texas. 
BURNS,  Edward  J.  (M  1923),  4716  Aldrich  Ave. 

S.,  Minneapolis,  Minn. 
BURNS,  John  R.  (J  1935;  5  1933),  (for  mail),  5035 

Forbes  St.,  Pittsburgh,  Pa.,  and  504  N.  Main  St., 

Wallingford,  Conn. 
BURNS,  Robert  (M  1934),  Engr.  (for  mail),  Coal 

Stoker  Sales   Co.,   500   N.   Craig  St.,   and  482 

Antenor  Ave.,  Pittsburgh,  Pa. 


BURRITT,  Charles  G,  (.4  1916),  Mgr.,  Minne- 
apolis Office  (for  mail),  Johnson  Service  Co.,  922 
Second  Ave.  S.,  and  Buckingham  Hotel,  Minne- 


apolis, Minn. 

BUSHNELL,  Carl  D.  (A  1921),  Pres.  (for  mail), 
The  Bushnell  Machinery  Co.,  1501  Grant  Bldg., 
Pittsburgh,  and  94  Pilgrim  Rd.,  Rosslyn  Farms, 
Carnegie,  Pa. 

BUTLER,  Peter  D.  (M  1922),  Salesman,  U.  S. 
Radiator  Corp.,  370  Lexington  Ave.,  New  York, 
N.  Y.,  and  (for  mail),  127  Edgewater  Rd.,  Grant- 
wood,  N.  J. 

BUTT,  Roderick  E.  W.  (J  1930),  Partner,  Crerar, 
Butt  &  Co.,  14  Regent  St.,  London  S.W.I.,  and 
(for  mail),  3  Orme  Court,  London  W2,  England. 

BUTTS,  Robert  L.  (S  1935),  64th  and  Norman- 
dale  Sts,,  Minneapolis,  Minn. 


CALDWELL,  Arthur  C.  (M  1930),  Estimator  and 

Engr.,  P.  Gormly  Co.,  155  North  10th  St.,  and 

(for  mail),  550  South  48th  St.,  Philadelphia,  Pa. 
CALEB,    David    (M    1923),    Engr.    (for    mail), 

Kansas  City  Power  &  Light  Co.,  1330  Baltimore 

Ave.,  and  141  Spruce  St.,  Kansas  City,  Mo. 
CALLAGHAN,  Philip  F.,  Jr.    (J  1929),    Sales 

Mgr.,  D.  G.  C.  Trap  &  Valve  Co.,  9  East  46th 

St.,  New  York,  and  (for  mail),  3003  Ave.   I, 

Brooklyn,  N,  Y. 
CALLAHAN,  Peter  J.  (M  1934),  Sr.  Draftsman, 

College  City  of  New  York  Project,  c/0  C.  B. 

Heweker,  Village  Hall,  Stapleton»  and  (for  mail), 

4057  Amboy  Rd.,  Great  Kills,  Staten  Island, 

N.  Y. 
CAMPBELL,  Alfred  Q.,  Jr.  (J  1933),  Sales  Mgr., 

E.   K.   Campbell   Cos.,   and   (for   mail),    1083 

Meriwether  Ave.,  Memphis,  Tenn. 
CAMPBELL,  Everett  K.*  (M  1920),  (Council, 

1931-1933),  Pres.  and  Treas.  (for  mail),  E.  K. 

Campbell  Heating  Co..  2445  Charlotte  St.,  and 

3717  Harrison  Blvd.,  Kansas  City,  Mo. 
CAMPBELL,  E.  K,,  Jr.  (J  1930),  Thermidaire 

Corp.,  2445  Charlotte  St.,  Kansas  City,  Mo. 
CAMPBELL,  F.  B.  (A  1927),  (for  mail),  American 

Radiator  Co.,  40  West  40th  St.,  New  York,  and 

245  Macon  St.,  Brooklyn,  N.  Y. 
CAMPBELL,  Robert  B.  (S  1934),  c/o  Mra.  L. 

Winn,  781  Ocean  Ave.,  Brooklyn,  N.  Y. 
CAMPBELL,  Thomas  F.  (M  1928),  MmneapoHs- 

Honeywell    Regulator    Co.,    1013    Penn    Ave., 

Wilklnaburg,  Pa. 
CANDEE,  Bertram  C.  (M  1933),  Partner,  Beman 

&  Candee,  374  Delaware  Ave.,  Buffalo,  and  (for 

mail),  19  Tremont  Ave.,  Kenmore,  N.  Y. 
CANNON,  C.  Newton  (/  1935;  5  1933),  General 

Electric  Co,,  and  (for  mail),  1104  Wendell  Ave., 

Schenectady,  N,  Y, 
CAREY,   James  A,    (M  1928),  Carrier  Engrg. 

Corp.,  Newark,  N.  J.,  and  (for  mall),  Vlllanova, 

Pa. 
CAREY,  Paul  C.  (M  1930),  (for  mall),  Runyon  & 

Carey,  33  Fulton  St.,  Newark,  and  31  Clare- 

mont  Dr.i  Maplewood,  N,  J, 
CARLE,  William  E,  (U  1926),  Pres.  (for  mail), 

Carle- Boehling  Co,,  Inc.,  1641  W,  Broad  St.,  and 

2220  Floyd  Ave»f  Richmond,  Va* 
CARLSON,  Everett  E.  (M  1932$  A  1929),  Br* 

Mgr,  (for  mall),  The  Powers  Regulator  Co.,  1010 

lx>uderoian  Bldg.,  and  66$2  Washington  Ave., 

St.  Louia,  Mo. 

CARMAN,  G«orft©  O.  (A  1031;  J  1928),  Lewis 
Institute,  Chicago,  III, 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


CARPENTER,  R.  H.  (M  1921),  (Council,  1930- 
1934),  Mgr.,  New  York  Office  (for  mail),  Nash 
Engrg.  Co.,  Graybar  Bldg.,  420  Lexington  Aye., 
New  York,  and  20  Jefferson  Ave.,  White  Plains, 
N.  Y. 

CARR,  Maurice  t.  (M  1931),  Director  (for  mail), 
Pittsburgh  Testing  Lab.,  P.  O.  Box  1646,  and 
Webster  Hall,  Pittsburgh,  Pa. 

CARRIER,  Earl  G.  (J  1929),  Estimating  Engr., 
Carrier  South  Africa  (Pty),  Ltd.,  20  Beresford 
House,  Simmonds  St.,  Johannesburg,  Transvaal, 
Union  of  South  Africa. 

CARRIER,  Willis  H.*  (M  1913),  (Presidential 
Member),  (Pres.,  1931;  1st  Vice-Pres.,  1930;  2nd 
Vice-Pres.,  1929;  Council,  1923-32),  Chairman 
of  the  Board  (for  mail),  Carrier  Corp.,  850 
Frelinghuysen  Ave.,  Newark,  and  Rensselaer 
Rd.,  Essex  Fells,  N.  J. 

CARTER,  Doctor  (M  1934),  Consulting  Engr., 
273  Ave.  Haig.,  Shanghai,  China. 

CARY,  Edward  B.  (M  1935),  Vice-Pres.,  John 
Paul  Jones,  Cary  &  Miller,  Inc.,  Cleveland,  and 
(for  mail),  3549  Daleford  Rd.,  Shaker  Heights, 
Ohio. 

CASE,  Walter  G.  (A  1930),  Tech.  Mgr.,  Ideal 
Boilers  &  Radiators,  Ltd.,  Ideal  House  Gt.,  Marl- 
borough  St.,  London  W.I.,  and  (for  mail),  66 
The  Ridgeway-Kenton,  Middlesex,  England. 

CASEY,  Byron  L.  (M  1921),  Sales  Engr.  (for 
mail),  Ilg  Electric  Vtg.  Co.,  182  N.  LaSalle  St., 
Chicago,  and  515  N.  Park  Ave.,  Park  Ridge,  111. 

CASEY,  Huntley  F.  (M  1931),  Sales  Repr.,  P.  O. 
Box  271,  E.  Falls  Church,  Va.,  and  (for  mail), 
756  E.  River  St.,  Anderson,  S.  C. 

CASH,  Tidie  T.  (A  1925),  Mgr.  (for  mail), 
Grinnell  Co.,  Inc.,  240  Seventh  Ave.  S.,  and  C17 
Kenwood  Pkwy.,  Minneapolis,  Minn. 

CASPERD,  Henry  W.  H.  (J  1930),  Engr.,  Carrier 
Engrg.  Co.,  Ltd.,  12  Mission  Row,  Calcutta, 
India,  and  (for  mail),  21  Robin  Hood  Lane, 
Sutton,  Surrey,  England. 

CASSELL,  John  D.*  (Life  Member;  M  1913), 
(Council,  1930-34),  2008  Walnut  St.,  Phila- 
delphia, Pa. 

CHANDLER,  Clark  W.  (J  1935;  S  1930),  (for 
mail),  Chandler  Co.,  and  1815  Ridgewood 
Terrace,  Cedar  Rapids,  Iowa, 

CHAPIN,  C.  Graham  (M  1933),  231  State  St.,  . 
New  London,  Conn. 

CHAPPELL,  Henry  D,  (M  1931),  Dist,  Mgr., 
V-Belt  Drive  Co.,  100  Morgan  Bldg.,  and  (for 
mail),  8019  Third  Ave.,  Detroit,  Mich. 

CHARLES,  Thomas  J.  (M  1934),  Pres.  (for  mail), 
Metropolitan  Air  Cond.  Corp.,  432  Fourth  Ave., 
New  York,  and  175  Marine  Ave.,  Brooklyn, 
N.  Y. 

CHARLET,  Louis  W.  (M  1934),  Mgr.,  New  York 
Br,  (for  mail),  Kewanee  Boiler  Corp.,  35-37 
West  39th  St.  New  York,  and  427  JRJch  Ave., 
Mt.  Vernon,  N,  Y. 

CHARLTON,  John  Felder  (A  1932),  Engr., 
Appraiser  (for  mail),  Box  2087,  and  1631  N.E. 
Fifth  St.,  Ft.  Lauderdale,  Fla. 

CHASE,  Chmincey  L,  (M  1931),  Ht£.  and  Vtg. 
Engr.,  Edward  E.  Ashley,  Cons.  Engr.,  10  East 
40th  St.,  New  York,  and  (for  mail),  8829  Ft. 
Hamilton  Pkwy.,  Brooklyn,  N,  Y. 

CHASE,  L.  Richard  (J  1931),  Br.  Mgr.  (for  mail), 
Buffalo  Forge  Co.,  &15  Dwight  Bldg.*  and  4822 
Wornall  Rd.,  Kansas  City,  Mo. 

CHEESEMAN,  Evans  W.  (5  1034),  Carnegie 
last,  of  Tech.,  Pittsburgh,  Pa. 

CHERNE,  Realto  E»  (J  1929),  Engr.,  Carrier 
Engrg,  Corp..  Chrysler  Bldg..  New  York.  N,  Y., 
and  for  mail),  126  DeHart  PI.,  Etabeth,  N.  J, 

CHERRY,  Lester  A.*  (M  1921),  Consulting  Engr. 
(for  mail),  Industrial  Planning  Corp.,  271 
Delaware  Ave.,  Buffalo,  and  155  Euclid  Ave., 
Kenmore,  Eric  Co*.  N.  Y* 

CHERVBN,  Victor  W.  (M  1928;  A  1920),  Chief 
Engr,  (for  mail),  Holland  Furnace  Co.,  and  326 
Maple  Ave.,  Holland,  Mich. 

CHESTER,  Thomaa^JM  1917},  Consulting  Engr., 
949  Chicago  Blvd.,  Detroit,  Mich. 

CHBSNUTT,  N,  P.  (S  1984),  760  De  Barr, 
Norman,  Okla. 


CHEYNEY,  Charles  C.  (.4  1913),  Asst.  Sales 
Mgr.  (for  mail),  Buffalo  Forge  Co.,  490  Broad- 
way, and  255  Lincoln  Pkwy.,  Buffalo,  N.  Y. 

CHIPPERFIELD,  W.  H.  (A  1934),  Service  Engr., 
Walker-Crosweller  Co.,  Ltd.,  20  Queen  Elizabeth 
St.,  S.E.I.,  and  (for  mail),  54  Lankers  Dr.,  N. 
Harrow,  Middlesex,  England. 

CHOFFIN,  C.  C.  (M  1919),  Pres-Treas.  (for  mail), 
The  Scholl-ChoSin  Co.,  Mahoning  Ave.  and 
Hogue  St.,  and  560  Tod  Lane,  Youngstown, 
Ohio. 

CHRISTENSON,  Harry  (A  1931),  Supt.  of  Htg. 
(for  mail),  Hunter  Prell  Co.,  311  Elm  St.,  and  85 
Wentworth  Ave.,  Battle  Creek,  Mich. 

CHRISTIAN,  Charles  W.  (Life  Member;  M  1913), 
Mgr.  (for  mail),  Chas.  W.  Christian  Co.,  P.  0. 
Box  292,  Charlotte,  and  1101  Providence  Rd., 
Myers  Park,  N.  C. 

CHRISTIE,  Alfred  Y.  (A  1933),  Salesman,  U.  S. 
Radiator  Corp.,  233  Vassar  St.,  Cambridge,  and 
(for  mail),  715  LaGrange  St.,  West  Roxbury, 

CHRISTMAN,  William  F.  (.4  1932;  /  1931),  (for 
mail),  Kroeschell  Engrg.  Co.,  2306  N.  Knox  Ave,, 
and  3912  N.  Hoyne  Ave.,  Chicago,  111. 

CHURCH,  Herbert  John  (M  1922),  Mgr.  (for 
mail),  Darling  Brothers,  Ltd,,  137  Wellington 
St.  W.,  Room  902  Toronto,  and  358  Main  St.  N., 
Weston,  Ont.,  Canada. 

CLARE,  Fulton  Warren  (M  1927),  Owner  (for 
mail),  Clare  &  CoM  120  Spring  St.  N.W.,  and 
935  Plymouth  Rd.,  Atlanta,  Ga. 

CLARKE,  Samuel  S.  (Life  Member;  A/  1900), 
Pres.  and  Mgr.  (for  mail),  S.  S.  Clarke  &  Co., 
Ltd.,  605  W.  Second  St.,  and  003  W.  Second  St., 
Calgary  Alberta,  Canada. 

CLARKSON,  Robert  C.,  Jr.  (M  1921),  6050 
O  verb  rook  Ave,,  Philadelphia,  Pa. 

CLARKSON,  W.  B.  (Life  Member,-  M  1919),  251 
Broadway,  Owatonna,  Minn. 

CLEGG,  Carl  (M  1922),  Dist.  Mgr.  (for  mail), 
American  Blower  Corp.,  311  Mutual  Bldg.,  and 
3321  Gillham  Rd.,  Kansas  City,  Mo. 

CLEGG,  Robert  R.  (A  1033),  Zone  Repr.,  Owens 
Illinois  Glass  Co.,  Industrial  Div,»  Lanclreth 
Bldg.,  and  (for  mail),  4515  Lindell  Blvd.,  St. 
Louis,  Mo. 

CLODFELTER,  John  L.  (A  1932),  Supt.  (for 
mail),  Carolina  Sheet  Metal  Corp.,  4210  Sansom 
St.,  Philadelphia,  and  West  Cheater  Pike  and 
Brief  Ave.,  Elizabeth  Manor  Apt.,  Upper  Darby, 

CLOSE,  Paul  D.*  (M  1928),  Chief  Engr.,  In- 
dustrial Uses  Div.  (for  mail),  Celotex  Co.,  919 
N.  Michigan  Ave.,  Chicago,  and  4622  Grove, 
Niles  Center,  III. 

CLOUGH,  Leslie  (M  1922),  Consulting  Engr.  (for 
mail),  Box  34  and  203  Pierce  Rd.,  Weymouth, 
Mass. 

CQCHRAN,  Lex  H.  (M  1934),  Dist.  Mgr.  (for 
mail).  American  Blower  Corp.,  Rialto  Bldg.,  and 
130  Camino  Del  Mar,  San  Francisco,  Calif, 

COB,  Ralph  T.  (M  1917),  Prop,  (for  mail),  The 
R,  T.  Coe  Cos.,  400  Reynolds  Arcade,  and  235 
Chile  Ave.,  Rochester,  N.  Y. 

COHAGBN,  Chandler  C.  (M  1919),  P»  <X  Box 
2100,  Billings,  Mont, 

COHEN,  Nftthan  (J  1935 j  S  1033),  2305  Loring 
PL,  New  York,  N,  Y* 

COHEN,  Philip  (M  1932),  DM.  Mgr,  (for  mail), 
B.  F.  Sturtevant  Co.,  407  E.  Ohio  Gae  Bldg., 
Cleveland  and  3681  Lynrtfield  Rd.,  Sh&ker 
Heights,  Ohio, 

COLBY,  Clyde  W.  (M  1016),  Coatulttaf  Bngr* 
(for  mail).  Old  School  House,  South  Hadley, 
Mass.,  and  4,0  Rosemere  Ave.,  Rye,  N,  Y, 

COLCLOUGH,  O,  T.  (A  1933),  Custodian, 
American  Leg&tlon.  American  Government 
Bldg.,  and  (for  mail),  407  Elgin  SL,  Ottawa 
Canada. 

COLE,  Edvrtta  Q.  (M  1931),  S82  Lebanon  St., 
Mel  rose,  Mass. 

COLE,  Grant  E.  (-4  1025),  489  Kittg  St.  W,, 
Toronto,  Ont.»  Canada* 


ROLL  OF  MEMBERSHIP 


COLEMAN,  John  B.  (M  1920),  Chief  Engr.  (for 
mail),  Grinnell  Co.,  Inc.,  275  W.  Exchange  St., 
and  237  Cole  Ave.,  Providence,  R.  I. 

COLLAMORE,  Ralph  (M  1904),  (Board  of 
Governor,  1913),  Secy.,  Smith,  Hinchman  & 
Grylls,  800  Marquette  Bldg.,  and  (for  mail), 
679  Pingree  Ave.,  Detroit,  Mich. 

COLLIER,  William  I.  (M  1921),  W.  I.  Collier  & 
Co.,  522  Park  Ave.,  Baltimore,  Md. 

COLLINS,  John  F.  S.,  Jr.  (M  1933),  Supervisor 
of  Steam  Utilization  (for  mail),  Alleghany 
County  Steam  Htg.  Co.,  Philadelphia  Co.  Bldg., 
435  Sixth  Ave.,  and  827  N.  Euclid  Ave.,  Pitts- 
burgh, Pa. 

COMSTOCK,  Glen  Moore  (A  1920),  Dist.  Repr. 
(for  mail),  L.  J.  Wing  Mfg.  Co.,  004  Chamber  of 
Commerce  Bldg.,  Pittsburgh,  and  154  College 
Ave.,  Beaver,  Pa. 

CONNELL,  Richard  F.  (M  1916),  Mgr.,  Capitol 
Testing  Lab.,  U.  S.  Radiator  Corp.,  1056  First 
National  Bank  Bldg.,  Detroit,  Mich. 

CONNER,  Raymond  M.  (M  1931),  Director  (for 
mail),  American  Gas  Assn.,  1032  East  62nd  St., 
and  271  East  216th  St.,  Cleveland,  Ohio. 

COOK,  Alton  B.  (S  1934),  533  S.  Flood,  Norman, 
Okla. 

COOK,  Benjamin  F.  (M  1920),  Consulting  Engr. 
(for  mail),  114  West  10th  St.  Bldg.,  Kansas  City, 
and  1720  Overtoil  Ave.,  Independence,  Mo. 

COOK,  Howard  A.  (A  1933),  Supt.,  Htg.,  Vtg. 
and  Sprinkling  (for  mail),  University  Plbg.  & 
Htg,  Co.,  3939  University  Way,  and  1433-33rd 
Ave.,  Seattle,  Wash. 

COOK,  Ralph  P.  (M  1930),  Engr.  of  Mech. 
Equip,  (for  mail),  Eastman  Kodak  Co.,  Kodak 
Park,  and  105  Falleson  Rd.,  Rochester,  N.  Y. 

COOMBE,  James  (A  1932).  Vice-Pres.  (for  mail), 
The  Wrn.  Powell  Co.,  2525  Spring  Grove  Ave., 
and  23«3  Grandin  Rd.,  Cincinnati,  Ohio. 

COON,  Thurlow  E.  (M  1910),  Pres.  (for  mail), 
The  Coon-De  Visser  Co.,  2051  W.  Lafayette,  and 
820  Edison  Ave..  Detroit,  Mich. 

COOPER,  Frederick  IX  (A  1930),  Sales  Engr., 
905  Holdcn  Ave.,  and  (for  mail),  1746  Longfellow 
Ave.,  Detroit,  Mich. 

COOPER,  John  W.  (M  1932;  A  1925;  J  1921), 
Repr.  (for  mail),  Buffalo  Forge  Co,,  1596  Arcade 
Bldg.,  St.  Louis,  and  312  E.  Big  Bend  Rd., 
Webster  Groves,  Mo. 

COPPERXJD,  Edmund  R.  (/  1933),  Asst.  Mgr. 
(for  mail),  Minneapolis  Plbg.  Co.,  1420  Nicollet 
Ave.,  and  4110  Nicollet  Ave.»  Minneapolis,  Minn. 

CORNELL,  J,  Clarence  (A  1930),  Checker 
(Mechanical),  12  South  12th  St.,  and  (for  mail), 
2823  W.  Allegheny  Ave.,  Philadelphia,  Pa. 

CORNWALL,  Georfte  I,  (M  1919),  Mgr.,  Boiler 
Dept.  (for  mail),  Hitchings  &  Co.,  701  Spring 
St.,  and  633  Madison  Ave.,  Elizabeth,  N.  J. 

CORRAO,  Joseph  (J  1933),  Engr.,  C,  C,  Moore 
Co.,  450  Mission  St.,  and  (for  mail),  85«lat 
Ave.,  San  Francisco,  Calif, 

CQRRIGAN,  James  A.  (J  1935;  5  1930),  2501  W. 
St.  Louis  Ave,,  St.  Louis,  Mo. 

COWARD,  Herbert  (M  1921),  Carrier  Bngrg, 
Corp.,  004  Washington  Bid*.,  Washington,  I>,  C. 

COX,  Harrison  F.  (A  1930),  243  Carroll  St., 
Pateraon,  N.  J. 

COX,  William  W.  Of  1923),  (for  mail).  Heating 
Service  Co,,  820  Columbia  St.,  and  6232-Slat 
Ave.  N.B.,  Seattle,  Wash, 

CRANSTON,  William  E.,  Jr.  (M  1931),  (Loa 
Angeles  Board  of  Governors,  1933-34),  Vice- 
Pres,  (for  mail).  Thermador  Electrical  Mfg.  Co., 
110  Llewellyn  St.  LOB  Angeles,  and  1912  Meri- 
dian Ave.,  South  Pasadena,  Calif. 

CRAWFORD:  John  H»,  Jr»  (J  1930),  372  High- 
land Ave*,  Orange,  N,  J. 

CRESSY,  Ralph  E.  (/  1020),  Sales  BngrM  Hoff- 
man Specialty  Co.,  500  Fifth  Ave.,  New  York, 
and  (tor  mail),  408  St,  Lawrence  Ave,,  BuSalo, 

GIUQUX,  Albert  A,*  (M  1019),  Chief  Engr.,  Htg. 
and  Vtg,  Dept,  Buffalo  Forge  Co.,  490  Broad- 
way* and  (for  mail),  250  Blaine  Ave,,  Buffalo, 

N,  Y. 


CRONE,  Charles  E.,  Jr.  (M  1922),  Secy-Treas. 
(for  mail),  Wendt  &  Crone  Co.,  2124  Southport 
Ave.,  and  1320  N.  State  St.,  Chicago,  111. 

CRONE,  Thomas  E.  (Life  Member;  M  1920), 
Salesman,  W.  A.  Russell  &  Co.,  Grand  Central 
Term.  Bldg.,  and  (for  mail),  542  West  112th  St., 
Apt.  10A,  New  York,  N.  Y. 

CROSS,  Robert  E.  (A  1931),  95  State  St.,  Spring- 
field, Mass. 

CUCCI,  Victor  J.  (M  1930),  Consulting  Engr., 
347  Madison  Ave.,  New  York,  N.  Y. 

CULBERT,  William  P.  (A  1929),  Secy,  (for  mail), 
Culbert-Whitby  Co.,  Inc.,  2019  Rittenhouse  St., 
Philadelphia,  and  929  Alexander  Ave.,  Drexel 
Hill,  Pa. 

GUMMING,  Robert  W.  (M  1928),  Mech.  and 
Sales  Engr.,  Sarco  Co.,  Inc.,  183  Madison  Ave., 
New  York,  and  (for  mail),  81  Alkamont  Ave., 
Scarsdale,  N.  Y. 

CUMMINGS,  Carl  H.  (A  1927;  J  1926),  Mgr.  (for 
mail),  Industrial  Appliance  Co.  of  New  England, 
250  Stuart  St.,  Boston,  and  41  Edgehill  Rd., 
Chestnut  Hill,  Mass. 

CUMMINGS,  G.  J.  (M  1923),  2001  Hoover  Ave., 
Oakland,  Calif. 

CUMMINS,  George  H.  (M  1919),  Dist.  Mgr.  (for 
mail),  Aerofin  Corp.,  616  United  Artist's  Bldg., 
and  17376  Wisconsin  Ave.,  Detroit,  Mich. 

CUNNINGHAM,  Thomas  M.  (M  1931;  A  1931; 
J  1930),  Production  Mgr.,  Carrier  Engrg.  Corp., 
180  N.  Michigan  Ave.,  Chicago,  111. 

CURRIER,  Charles  H.  (M  1919),  Vice-Pres.  (for 
mail),  Ross  Heater  &  Mfg.  Co.,  Inc.,  1407  West 
Ave.,  and  Park  Lane  Apts.,  33  Gates  Circle, 
Buffalo,  N.  Y. 

CURTIS,  Herbert  F.  (A  1934),  Berea,  Ohio. 

CUSHMAN,  Lester  D.  (M  1930),  89  Traincroft 
St.,  Medford,  Mass. 

CUTLER,  Joseph  A.  (M  1916),  (Council,  1917- 
1926),  Vice-Pres.  (for  mail),  Johnson  Service  Co., 
1355  Washington  Blvd.,  Chicago,  and  649 
Hinman  Ave.,  Evanston,  111. 

D 

DAHLSTROM,  Godfrey  A.  (A  1927),  Htg.  Sales 
Engr.,  Central  Supply  Co.,  312  S.  Third  St.,  and 
(for  mail),  3721-47th  Ave.  S.,  Minneapolis,  Minn. 

DAILEY,  James  A.  (A  1920),  31-64-30th  St., 
Astoria,  L.  I.,  N,  Y. 

DAKJN,  Harold  W.  (J  1934),  Asst.  Engr,,  Wagner 
Engrg.  Corp.,  22  Dunham  St.,  Pittsfield,  and  (for 
mail),  169  Park  Ave.,  Dalton,  Mass. 

DALLA  VALLE,  J.  M.*  (J  1933),  Asst.  Sanitary 
Engr.,  U.  S,  Public  Health  Service,  19th  and 
Constitution  Ave..  Washington,  D.  C.,  and  (for 
mail),  17  Jones  Bridge  Rd.,  Chevy  Chase,  Md. 
'  DALY,  Charles  P.  (A  1935),  Contractor  (for  mail), 
Rantmann  Plbg.  &"Htg.  Co.,  115  Jackson  St., 
and  2438  Queen  Anne.  Seattle,  Wash. 

DALY,  Robert  E.  (M  1931),  Executive  Dept.  (for 
mail),  American  Radiator  Co.,  40  West  40th  St., 
and  12  East  88th  St.,  New  York,  N.  Y. 

DAMBLY,  A.  Ernest  (M  1924;  J  1921),  (for  mail), 
H.  B.  Hackett,  901  Architects  Bldg.,  Phila- 
delphia, Pa.,  and  Harvey  Cedars,  N.  J. 

DANFORTH*  N.  Lorlnfc  (M  1919),  John  W. 
Danforth  Co,,  72  Elllcott  St.  Buffalo,  N.  Y. 

DARBY,  Marion  H.  (J  1930),  Sales  Engr.  (for 
mail),  Carrier-Brunswick  de  Mexico,  S,A., 
Edincio  Cidoaa  Deapacho,  101,  Uruguay  55, 
Mexico,  D.F.,  Mexico. 

DARLING,  Arthur  B.  (A  1929),  Asst.  Sales 
Mgr*  (for  mail).  Darling  JBros,,  Ltd-,  140  Prince 
St.,  and  4216  Dorchester  St.  WM  Montreal,  P.  Q,, 
Canada. 

DARLINGTON,  Allan  P.  (M  1930),  Salesman 
(for  mail),  American  Blower  Corp.,  2539  Wood- 
ward Ave..  and  3605  Devonshire,  Detroit,  Mich. 

DARTS,  John  A.  (M  1919),  Kewanee  Boiler  Co,, 
Inc.,  570  Seventh  Ave.,  New  York,  N,  Y. 

DAUCH,  Emll  O.  (M  1921),  Secy-Treas.  (for 
mall),  McCormlck  Plbg*  Supply  Co..  1075 
Bailey  Ave.,  and  The  Whittler  Hotel,  Detroit, 
Mich. 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


DAVENPORT,  R.  F.  (A  1933),  77  James  St.  E., 

Brockville,  Ont.,  Canada. 
DAVIDSON,  L.  Clifford  (M  1927),  Associate  Dist. 

Mgr.  (for  mail),  Buffalo  Forge  Co.,  220  South 

16th  St.,  and  6312  Sherwood  Rd.,  Philadelphia, 

Pa. 
DAVIDSON,  Philip  L.  (M  1921;  J  1921),  Asst. 

Dist,  Mgr.  (for  mail),  Carrier  Engrg.  Corp. ,'12 

South   12th  St.,   Philadelphia,  and   14  Radnor 

Way,  Radnor,  Pa. 
DA  VIES,  George  William  (M  1918),  Managing 

Dir.  (for  mail),  Htg.,  Vtg.  &  Domestic  Engrs., 

79  Maclaggan  St.,  Dunedin,  and  145  Kenmure 

Rd.,  Mornington,  New  Zealand. 
DAVIS,   Arthur  C.*   (M  1920),  Supt.  of  Main- 
tenance, The  Port  of  New  York  Authority,  111 

Eighth  Ave.,  New  York,  N.  Y.,  and  (for  mail), 

73  Preston  St.,  Ridgefield  Park,  N.  J. 
DAVIS,  Arthur  Folsom  (M  1934"),  Vice-Pres.  (for 

mail),  The  Johnson  &  Davis  Plbg.  &  Htg.  Co., 

2235    Arapahoe    St.,    and    1901    Ivanhoe    St., 

Denver,  Colo. 
DAVIS,  Bert  C.  (JW  1904),  (Council,  1917),  Pres. 

and   Treas.    (for  mail),   American  Warming  & 

Ventilating  Co.,  317-19  Pennsylvania  Ave.,  and 

003  W,  Church  St.,  Elmira,  N.  Y. 
DAVIS,  Calvin  R.  (jtf  1927),  Br,  Mgr.  (for  mail), 

Johnson  Service  Co.,  2328  Locust  St.,  and  7534 

Westmoreland  Dr.,  St.  Louis,  Mo. 
DAVIS,  James  R.  (S  1034),  2111  Abington  Rd., 

and  (for  mail),  9704  Miles  Ave.,  Cleveland,  Ohio. 
DAVIS,  Joseph  (M  1927;  A  1920),  Owner,  Htg. 

Engr.  and  Contractor  (for  mail),  007  Root  Bldg., 

70   W.    Chippewa,   and    llJti   Huntington  Ave., 

Buffalo,  N.  Y. 
DAVIS,  Otis  E.  (M  1929;  A  1925),  1501  Fourth 

Ave.,  Scotts  Bluff,  Nebr. 
DAVIS,  Rowland  G.  (A  1921),  Sales  Repr.,  887 

Nela  View  Rd.,  Cleveland  Heights,  Ohio. 
DAVISON,    Robert  L.    (M    1934),    Director   of 

Research  (for  mail),  John  B,  Pierce  Foundation, 

40  West  40th  St.,  and  2S  East  10th  St.,  New 

York,  N,  Y. 
DAWSON,  Eugene  F.  (M  1034),  Asst.  Prof.  Mcch. 

Engr.  (for  mail),  University  of  Oklahoma,  and 

910'  S,  mood  St.,  Norman,  Okla. 
DAWSON,  Thomas  L.  (M  1930),  Pres,  (for  mail), 

Thomas  L,  Dawson  Co.,  2035  Washington  St., 

Kansas    City,    Mo.,    aud    f)6th   and    Shawnee 

Mission  Rd.,   Rosedale  Station,   Kansas  City, 

Kans. 
DAY,    Harold    C.    (A    1934),    Mgr.,    American 

Radiator  Co.,  374  Delaware  Ave.,  and  (for  mail), 

Buffalo  Athletic  Club,  Delaware  Ave.,  Buffalo, 

N.  Y. 
DAY,  V.  S.*  (M  11)24) ,  En«n  tfor  mail),  Carrier 

Kngrg.  Corp.,  850  Krellnglmysen  Ave.,  Newark, 

and  100  Summit  Ave.,  Summit.  N,  J. 
DEAN,  Charles  L,  (M  1032) f  Asst,  Prof,  M«ch, 

Engrg..  University  of  Wisconsin,  and  (for  mail), 

2603  Stevens  St.,  Madison,  Wia, 
DEAN,  Ffank  J.,  Jr,  7s  19#4),  Clerical,  Automatic 

Electric  Co,,  1033  W,  Van  Buren,  and  (for  mail), 

12Q  S,  Central  Ave.,  Chicago,  111. 
DeBLOIS,  Lewis  A.  (M  1034),  Consulting  Engr,, 

485  East  57th  St.,  New  York,  N,  Y, 
DEELY,  James  J,  (J  1933),  Sales  Engr.,  Brooklyn 

Union  Gaa  Co.,  180  Remsen  St.,  and  (for  mail), 

Hotel  St.  George,  Brooklyn,  N,  Y. 
DeLAND,  Charles  W.  CM  1024;  J  1923),  Secy- 

Treas.  (for  mail),  C.  W.  Johnson  Co.,  Inc.,  211 
»N.  Desplaines  St.,  and  2021  Eatea  Ave,,  Chicago, 

DENNY,  Harold  R,  (A  1934),  Mgr.  Mereh,  Dept, 
American  Blower  Corp.,  401  Broadway,  New 
York,  N,  Y. 

DBUTCHMAN,  Julius  (/  1935;  S  1933),  1-3 
Welleiley  Ave,,  Yonkera,  N.  Y. 

DEWEY,  &.  P.  (M  1984),  Chief  Engr,  (for  mail), 
Barb«r~Colman  Co.,  and  2301  Oxford  St.. 
Rockford,  111. 

DIBBLE,  S.  E**  (M  1917),  (Presidential  Mamfow), 
(Pres,,  1925:  lat  Vice-Pres*.  1924;  2nd  Vice- 
£re».»  1922;  Council,  1921*1926).  Supt.,  Thomas 
Ran  ken  Patton  School*  BUssabethtown,  Pa. 


DICE,  Eugene  S.  (5  1933),  7141  Upland  St., 
Pittsburgh,  Pa. 

DICKENSON,  Frederick  R.  (A  1934),  Dist. 
Mgr.  (for  mail),  American  Blower  Corp.,  1302 
Swetland  Bldg.,  Cleveland,  and  3435  Menlo  Rd., 
Shaker  Heights,  Ohio. 

DICKEY,  Arthur  J.  (M  1921),  9  Mossom  Pi., 
Toronto,  Ont.,  Canada. 

DICKSON,  Robert  B,  (M  1919),  Pres.  (for  mail), 
Kewanee  Boiler  Co.,  Inc.,  Franklin  St.  and  Q 
Tracks,  and  409  E.  Prospect  St.,  Kewanee,  111. 

D'IMOR,  Elton  J.  (M  1933),  Br.  Mgr.  and  Engr., 
The  Trane  Co.,  LaCrosse,  Wis.,  and  (for  mail), 
ISO  N.  Auburndale  St.,  Apt.  7,  Memphis,  Tenn. 

DISNEY,  Melvin  A.  (A  1934),  Co-Partner, 
Mfr's.  Reprs.,  HtR.,  Vtg.,  and  Air  Cond.  Equip, 
(for  mail),  4301 H  Main  St.,  and  8024  Merrier, 
Kansas  City,  Mo. 

DISTEL,  Frank  (M  1918),  Owner,  Distel  Heating 
Equipment  Co.,  404-406  Kalamazoo  Plaza  (for 
mail),  P.  O.  Box  133,  and  1011  W.  Genesce  St., 
Lansing,  Mich. 

DIVER,  M.  L.  (M  1925),  Consulting  Engr.,  P.  O. 
Box  1016,  San  Antonio.  Texas. 

DIXON,  Arthur  G.  (A/  1928),  Sales  M«r.  (for 
mail),  Mocline  Mfg.  Co.,  and  442  Wolif  St., 
Racine,  Wis. 

DOBBS,  C.  E.  (A  1921),  Rcpr.,  Burnham  Boiler 
Corp,,  31st  and  Jefferson  Sta,t  Philadelphia,  Pa., 
and  (for  mail),  72  Berlin  Ave.,  Haddonticld,  N.  J. 

DODDS,  Forrest  F.  (AJ  1920),  Br.  Mgr.  (for  mail), 
American  Radiator  Co.,  1023  Grand  Ave.,  and 
235  Ward  Pkwy.,  Kansas  City,  Mo. 

DODGE,  Harry  G.  (A  1034),  Vice-Pres.,  Metro- 
politan Pipe  &  Supply  Co.,  14 fl  Broadway, 
Cambridge,  and  (for  mail),  28  Rustic  Rd., 
Melrose  Highland,  Mass. 

DOERING,  Frank  L.  (M  1010),  Salesman, 
American  Radiator  Co.,  210  Denver  Ave.r 

DOHERTY, 'Russell  (A  1920),  Chicago  Dist. 
M«r.  (for  mail),  National  Radiator  Corp,,  1111 
East  8Hrd  St.,  Chicago,  and  300  Forest  Ave., 
Oak  Park,  III. 

DOLAN,  Raymond  G.  {M  1920;  A  1920;  /  192ii), 
Secy-Treas.  (for  mail),  Tom  Dolan  Htg.  Co., 
Inc.,  614  W.  Grand,  and  2112  West  20th,  Okla- 
homa City,  Okla. 

DONNELLY,  James  A.*  (M  mm,  (Treasurer, 
1012-1914),  Urgent,  W.  Va. 

DONNELLY,  RuaseH  (M  1023),  Sales  Kngn  (for 
mall),  Nash  Engrg,  Co.,  Graybar  Bldg,,  420 
Lexington  Ave.,  New  York,  N.  V. 

DONOVAN,  William  J,  (A  1930),  2239  North 
27th  St.,  Philadelphia,  Pa, 

DONZELLI,  Enrico  (U  1933),  Piazza  SS  Pietro  e 
Lino,  No.  4.,  Milan,  Italy. 

DORFAN,  Morton  I.  (M  1020),  Mgr,  Du«t 
Collecting  l)iv,,  Blaw-Knox  Co,,  f>,  C),  Box  1108, 
and  (for  mail),  0357  Morrowfield  Ave,,  Pitts- 
burgh, Pa. 

DORNHEIM,  G.  A.  (M  1912;  J  1006),  15  Hamil- 
ton Ave,,  Brotixvilto,  N,  Y. 

DORSE  Y,  Francis  C.  (M  1020),  Engr,  and 
Contractor  (for  mall),  Fnmdti  C.  Doraey,  ine,,, 
4520  Schenley  Rd.,  Roland  Park,  and  212 
Gittings  Ave,,  Baltimore,  Md, 

DOSTBR,  Alexis  (A  1634),  sSecy.  (for  mail),  The 
Torringtori  Mfg,  Co.,  70  Franklin  S|,,  Torrington* 
and  Lltchfield,  Conn. 

DOUGHTY,  Charlos  Jofett  (M  W»fi)t  Prei,  and 
Managing  Director  (for  mail),  C.  J»  Doughty  & 
Co.,  F«d.  Inc..  U,  S.  A»f  30  Brenan  Rd,»  aad  1020 
Ave,  Joffre,  Shanghai*  China, 

DOVOLIS,  Nick  J.  (S  m&)>  S403  Chicago  Av«», 
Minneapolifl.  Minn. 

DOWNE,  Edward  R*  (M  1927)*  American  Ga* 
Products  Corp.,  40  Wett  40th  St,»  Nw  York, 
N.  Y. 

DOWNE,  Henry  S.  (L$f*  MMbtri  M  1805) ,  Cl« 
Natioaalc  dcs  Rtidiatoura,  140  Blvd.  Hausaman, 
ParifliFrance. 

0OWNKS,  Hftaty  H*  (M  1988) tMgr.  Navy ftl^Lp* 
Div.  (for  mall)»  American  Blower  Coit>*»  «*< 
Woodward  Bldg,,  Wa«Wn8*0%  D*  C.»  ana  460$ 
Davidson  0r.»  Chevy  Cfett«»  ]Sld» 


14 


ROLL  OF  MEMBERSHIP 


DOWNES,  Nate  W.  (M  1917),  (Council,  1928- 
1930),  Chief  Engr.  and  Supt.  of  Bldgs.  (for  mail), 
School  Dist.  of  Kansas  City,  317  Finance  Bldg., 
and  2119  East  68th  St.,  Kansas  City,  Mo. 

DOWNS,  Sewell  H.  (M  1931),  Chief  Engr., 
Clarage  Fan  Co.,  and  (for  mail),  211  Creston 
Ave.,  Kalamazoo,  Mich. 

DOYLE,  William  J.  (M  1920),  Factory  Mgr.,  The 
Williamson  Heater  Co.,  4558  Marburg  Ave.,  and 
(for  mail),  3766  Hyde  Park  Ave.,  Cincinnati, 
Ohio. 

DRINKER,  Philip*  (M  1922),  Assoc.  Prof,  (for 
mail),  Harvard  School  of  Public  Health,  55 
Shattuck  St.,  Boston,  and  Puddingstone  Lane, 
Newton  Center,  Mass. 

DRISCOLL,  William  H.*  (M  1904),  (Presidential 
Member),  (Pres.,  192G;  1st  Vice-Pres.,  192 A;  2nd 
Vice-Pres.,  1924;  Treas.,  1923;  Council,  1918- 
1927),  (for  mail),  Thompson-Starrett  Co.,  Inc., 
444  Madison  Ave.,  New  York,  N.  Y.,  and  50 
Glen  wood  Ave.,  Jersey  City,  N.  J. 

DuBOIS,  Louis  J.  (M  1931),  Air  Concl.  Engr., 
York  Ice  Machinery  Corp.,  117  South  llth  St., 
and  (for  mail),  7337a  Lindell  Ave.,  St.  Louis,  Mo. 

DUBRY,  Ernest  E.  (M  1924),  Asst.  Supt.,  Central 
Htg.,  The  Detroit  Edison  Co.,  2000  Second  Ave., 
and  (for  mail),  9116  Dexter  Blvd.,  Detroit,  Mich. 

DUDLEY,  William  Lyle  (M  1922),  Vice-Pres.  (for 
mail),  Western  Blower  Co.,  1800  Airport  Way, 
and  S14-32nd  Ave.,  Seattle,  Wash. 

DUFF,  Kennedy  (M  1915),  Mgr.  (for  mail), 
Johnson  Service  Co.,  28  East  29th  St.,  New 
York,  N,  Y.,  and  9  Park  Ave.,  Maplewood,  N.  J. 

DUG  AN,  Thomas  M.  (M  1920),  Sanitary  and 
Htg.  Exifir.,  National  Tube  Co.,  Fourth  Ave.  and 
Locust  St.,  and  (for  mail),  1308  Freemont  St., 
McKcesport.  Pa. 

DUGGER,  Earl  R.  (S  1934),  8409  Classen, 
Oklahoma  City,  Okla. 

DUNCAN,  George  W-,  Jr.  (M  1923),  2512 
Ben  venue  Ave.,  Berkeley,  Calif. 

DUNCAN,  James  R.  (M  1923),  Carrier  Austra- 
lasia, Ltd.,  5(5  Hunter  St.,  Sydney,  Australia. 

DUNCAN,  William  A.  (A  1930),  Dist.  Service 
Engr.  (for  mail),  Dominion  Oxygen  Co.,  Ltd.,  92 
Adelaide  St.  W,,  and  20  Tyrol!  Ave.,  Toronto, 
Out,,  Canada. 

DUNHAM,  Clayton  A.*  (M  1911),  Tres.  (for 
mail),  C.  A.  Dunham  Co.,  450  1C.  Ohio  St., 
Chicago,  arid  lf>0  Maple  Hill  Rd.,  Glcncoc,  111. 

DUKKEE,  Merritt  E,  (A  1930),  Sales  Engr.  (for 
mail),  C.  A.  Dunham  Co.,  101  Park  Ave.,  New 
York,  and  254  Martina  Ave,,  White  Plains,  N.  Y. 

BURNING,  Edward  H.  (J  1931),  Commercial 
Sales,  Dallas  Gas  Co.,  Harwood  and  Jackson 
Sta.,  and  (for  mail),  1880  Moaar  St.,  Dallas, 
Texas, 

DURYEA,  Albert  A.  (J  1935;  S  1933),  151  Belden 
Point,  City  Island,  N.  Y. 

DUSOSSOIT,  Edmond  A.  (M  1920),  Treas,  (for 
mail),  Lynch  &  Woodward,  inc..  320  Dover  St., 
Boston,  and  10  Hancock  Ave,,  Newton  Center, 
Mass, 

DWYER,  Thomas  F.  (M  1923),  Meeh*  Kngr.  (for 
mail),  Board  of  Education,  49  Flatbueh  Ave. 
Ext.,  Brooklyn,  and  1183  Clay  Ave,,  New  York, 
R  Y, 

DYER,  Qrvilta  1C.  (M  1919),  Mgr..  Blower  Div. 
(for  mail),  Buffalo  Forge  Co,,  490  Broadway,  sind 
11  Rusted  Ave.»  Buffalo,  N.  Y. 

E 

EA0IE,  John  G,  (M  1909),  Eadie,  Freund  £ 
Campbell  Co,,  110  West  40th  St.,  New  York, 
N.  VT 

EAGAR,  R,  Frank  (M.  1022),  98  Edward  St., 
Halifax,  N.S.,  Canada, 

EAICINS,  Walter  (M  1928),  830  E,  Phil  Eltena 
StM  Germantown,  Philadelphia,  Fa, 

EASTMAN,  Carl  B.  (M  1932;  A  1032;  J  1929), 
Mgr*  Philadelphia  Sake  Ofta,  C,  A.  Dunhajn 
Co.,  1500  walnut  St.,  Philadelphia,  and  (for 
mil),  7247  Calvin  Ed,.  Upper  Darby,  Pa,  > 


EASTWOOD,  E.  O.  (M  1921),  (Council,  1931- 
1934),  Prof,  of  Mech.  Engrg.  (for  mail),  Uni- 
versity of  Washington,  and  4702-12th  Ave.  N.E., 
Seattle,  Wash. 

EATON,  Byron  K.  (M  1920),  75  N.  Park  Rd., 
La  Grange,  III. 

EATON,  William  G.  M.  (A  1934),  Sales  Engr., 
Pease  Foundry  Co.,  Ltd.,  US  King  St.  E., 
Toronto  2,  and  (for  mail),  59  Symington  Ave., 
Toronto  9,  Canada. 

EBERT,  William  A.  (M  1920),  Mech.  Contractor 
(for  mail),  1026  W.  Ashby,  and  2151  W.  Kings 
Highway,  San  Antonio,  Texas. 

EDWARDS,  Daniel  F.  (M  1920),  2340-42  Pine 
St.,  St.  Louis,  Mo. 

EDWARDS,  Don  J.  (.4  1933),  Vice-Pies,  (for 
mail),  General  Heat  &  Appliance  Co.,  94  Massa- 
chusetts Ave.,  Boston,  and  40  Rockledge  Rd., 
Newton,  Mass. 

EDWARDS,  Paul  A.  (M  1919),  Pres.  (for  mail), 
The  G.  F.  Higgins  Co.,  60S  Wabash  Bldg.,  and 
3074  Pinehurst  Ave.,  Pittsburgh,  Pa. 

EELLS,  Henry  B.  (M  1926),  New  York  Mgr., 
Barnes  &  Jones,  Inc.,  101  Park  Ave.,  New  York, 
and  (for  mail),  1049  East  27th  St.,  Brooklyn, 
N.  Y. 

EGGLESTON,  Lewis  W.  (M  1921),  American 
Radiator  Co.,  5961  Lincoln  Ave.,  Detroit,  Mich. 

EGGLY,  Harry  J.,  Jr.  (M  1933),  Consulting 
Engr.  (for  mail),  1805  Walnut  St.,  Philadelphia, 
and  Elkins  Park  Apts.,  Elkins  Park,  Pa. 

EHRLICH,  M.  William*  (M  1916),  Chief  Engr., 
Commodore  Heaters  Corp.,  11  West  42nd  St., 
New  York,  N.  Y.,  and  (for  mail),  56  Ridge  Rd., 
Lyndhurat,  N.  J. 

EICHBERG,  W.  Roy  (M  1929),  Pres.  (for  mail), 
Carolina  Sheet  Metal  Corp.,  4210  Sansom  St., 
Philadelphia, 'and  828  Turner  Ave.,  Drexel  Hill, 
Pa. 

EICHER,  HuBert  G.  (M  1922),  State  Director, 
School  Bldgs.,  Div.  Dept.  of  Public  Instruction, 
State  Capitol,  and  (for  mail),  103  South  St., 
Harrisburg,  Pa. 

EISS,  Robert  M.  (M  1933;  ,4  1933;  J  1930),  (for 
mail),  72  Brantwood  Rd.,  c/o  Buffalo  N.  Y.  P.  O., 
Eggertsville,  N.  Y, 

ELLINGWOOD,  Elliott  L.  (M  1909),  354  S. 
Spring  St.,  Los  Angeles,  Calif. 

ELLIOT,  Edwin  (M  1929),  (for  mail),  Edwin 
Elliot  &  Co.,  560  North  16th  St.,  Philadelphia, 
and  403  W.  Price  St.,  Germantown,  Philadelphia, 
Pa. 

ELLIOTT,  Louis  (M  1932),  Consulting  Mech. 
Engr.,  Electric  Bond  &  Share  Co.,  2  Rector  St., 
Room  1914,  New  York,  N.  Y.  < 

ELLIOTT,  Norton  B.  &l  1934),  Br.  Mgr., 
American  Blower  Corp.,  1011  Majestic  Bldg., 
and  (for  mail),  5170  N.  Idlewild  Ave.,  Mil- 
waukee, Wis. 

ELLIS,  Ernest  E.  (M  1922),  Secy-Treas.,  F.  A. 
Ellis  &  Co.,  Inc.,  840  Center  St.,  Wirmetka,  III. 

ELLIS,  Frederick  E.  (M  1923),  Sales  Mgr.  (for 
mail),  Imperial  Iron  Corp.,  Ltd.,  30  Jefferson 
Ave,,  Toronto,  and  9  Princeton  Rd.,  Kingaway 
P,  O.  Toronto  3,  Ontario,  Canada. 

ELLIS,  Frederick  R.  (M  1913),  Sales  Engr., 
Buerkel  &  Co.,  Inc.,  18*24  XJnion  Park  St., 
Boston,  and  (for  mail),  131  Beacon  St.,  Hyde 
Park,  Mass. 

ELLIS,  Harry  W.  (M  1923;  A  1909),  Pres., 
,  Johnson  Service  Co.,  507  E.  Michigan  St., 
Milwaukee,  Wla, 

EMERSON,  Ralph  R.  (M  1922),  48  Gay  St., 
Newtonville,  Mass. 

EMERY,  Hugh  (J  103tf;  S  1933),  20  Morgan  PL, 
N.  Arlington,  N.  J. 

EMMERT.Lwther  D.  (M  1010),  Repr.  (for  mail), 
Buffalo  Forge  Co,,  Room  1909,  20  N.  Wacker 
Dr.»  Chicago,  and  1704  Hinman  Ave.,  Evanston, 
111, 

EMMONS,  Neat  L.  (S  1984),  1100  East  19th, 
Oklahoma  City,  Okla. 

ENGEL,  Edward  (J  1933),  Design  Draftsman  (for 
mail),  Hull  Div.,  U.  S.  Navy  Yard,  and  2608 
North  30th  St,,  Philadelphia,  Pa. 


15 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


ENGLE,  Alfred  (A  1923),  Sales  Mgr.  (for  mail), 
Jenkins  Bros.,  80  White  St.,  New  York,  and  1 
Edgewood  Rd.,  Scarsdale,  N.  Y. 

EPPLE,  Arnet  B.  (J  1934),  Student  Engr.,  B.  F. 
Sturtevant  Co.,  Boston,  and  (for  mail),  8  Elm  St., 
Hyde  Park,  Mass. 

EPPRIGHT,  John  O.  (J  1934),  3928  Benton 
Blvd.,  Kansas  City,  Mo. 

ERDLE,  Gardner  F.  (A  1933),  Mfrs.  Repr.  (for 
mail),  374  Delaware  Ave.,  Buffalo,  and  19 
Trernont  Ave.,  Kenmore,  N.  Y. 

ERICKSON,  Harry  H.  (.4  1929),  Sales  Engr., 
General  Fittings  Co.,  804  Architects  Bldg.,  and 
(for  mail),  5909  North  21st  St.,  Philadelphia,  Pa. 

ERICKSON,  Martin  E.  (A  1926),  Supt.,  Main- 
tenance, Board  of  Education,  and  (for  mail), 
1533  South  74th  St.,  West  Allis,  Wis. 

ERICSSON,  Eric  B.  (M  1933),  Engr.,  Custodian, 
Board  of  Education,  and  (for  mail),  605  West 
116th  St.,  Chicago,  111. 

EVANS,  C.  A.  (M  1919),  527  Massachusetts  Ave., 
Buffalo,  N.  Y. 

EVANS,  Edwin  C.  (M  1919),  Consulting  Engr., 
2953  Zephyr  Ave.,  Corliss  Station  P.  O.,  Pitts- 
burgh, Pa. 

EVANS,  William  A.  (M  1918),  Mfr.  and  Agent, 
W.  A.  Evans  &  Co.,  180  Broadway,  New  York, 
N.  Y.,  and  (for  mail),  24  Woodland  Rd.,  Maple- 
wood,  N.  J. 

EVELETH,  Charles  F.*  (M  1911),  2030  East 
115th  St.,  Cleveland,  Ohio. 

EVERETTS,  John,  Jr.*  (A  1935;  J  1929),  Engr. 
(for  mail),  W.  L.  Fleisher,  11  West  42nd  St., 
New  York,  and  S3  Kenihvorth  PL,  Brooklyn, 
N.  Y. 

EVLETH,  Everett  B.  (A  1927),  Div,  Mgr.  (for 
mail),  Minneapolis-Honeywell  Regulator  Co., 
43  E.  Ohio,  Chicago,  and  1023  Ashland  Ave., 
Wilmette,  111. 

F 

FABER,  Dr.  Oscar  (M  193-1),  Consulting  Engr. 
(for  mail),  Romney  House,  Marsham  St.,  West- 
minster, London  and  Hayes  Court,  Kenley, 
Surrey,  England, 

FAGIN,  Daniel  J.  (M  1032),  Heating  Engr., 
House  Htg,  Div.,  Laclede  Gas  Light  Co.»  llth 
and  Olive,  and  (for  mail),  4920  Chippewa  Ave., 
St,  Louis,  Mo. 

FAHNESTOCK,  Maurice  K.*  (M  1927),  Re- 
search Asst.  Prof,  (for  mail),  University  of 
Illinois,  214  M.  E.  Laboratory,  and  701  W, 
California  St.,  Urbana,  111. 

FAILE,  Edward  H*  (M  1934),  Consulting  ICngr,, 
E.  H.  Kaile  Co.,  44r  Lexington  Ave,,  Hew  York, 
N.  YM  and  (for  mail),  1336  Fairfxeld  Ave,, 
Bridgeport,  Conn. 

FALTENBACHBR,  Harry  J,  (,M  1930),  235  E. 
Wister  St.,  Philadelphia,  Pa. 

FALVEY,  John  0.  <M  1M)>  6686  Perahing  Ave,» 
University  City,  Mo. 

FAMILETTI,  A,  Robert  (J  1930)>  2812  Wtorton 
St.,  Philadelphia,  Pa. 

FANSLER,  P.  E.  (A  1927),  Editor  (for  mall),  Oil 
Heat,  167  Madison  Ave.,  New  York,  N,  Y..  and 
CatonsviUe,  Md> 

FARLEY.  W.  F,  (M  1030),  Saieaman,  American 
Radiator  Co.,  40  West  40th  St.,  New  York,  and 
(for  mail),  28  Kirn  St.,  New  Rocheile,  N.  Y, 

FARLEY,  wuiouahby  s.  (j  ioaa),  partner, 

Farley  &  Luther,  120  S.  Union  St.,  and  (for  mail), 
Sill  Montague  St.,  Danville,  Va, 

FARNHAM,  Roswell  (M  1020),  (Council,  1927- 
1932),  DIst.  Mgr,,  Knjprg,  Sales  ffor  mail), 
Buffalo  Forge  Co.,  P.  O.  Box  0S5,  ana  5  Claren- 
don PI.,  Buffalo,  N,  Y, 

FARNSWORTH,  John  G.  (J  1931),  Gas  House 
Htg,  Engr.  (for  mail),  Central  Illinois  Light  Co,, 
310  S,  Jefferson  St.,  and  313  Crescent  St.,  Peoria, 
111. 

FARRAR,  Cecil  W,  (M  1020;  A  1918),  (Treai,, 
1930;  Council,  1980),  Pres.  (for  mail),  Excelso 
Products  Corp.,  ISO/  Klniwood  Ave.»  and  29 
Oakland  PL,  Buffalo,  N,  Y, 


FAUST,  Frank  H.*  (J  1930),  Engr.,  Air  Cond. 

Dept.  (for  mail),  General  Electric  Co.,  1  River 

Rd.,  and  114  Union  St.,  Schnectady,  N.  Y. 
FAY,  Francis  G.  (M  1925),  Raisler  Heating  Co., 

129-31  Amsterdam  Ave.,  New  York,  N.  Y. 
FEBREY,   Ernest  J.    (M   1903),   Htg,  and   Air 

Cond.  (for  mail),  61G  New  York  Ave,  N.W.,  and 

2331  Cathedral  Ave.,  Washington,  D.  C. 
FEEHAN,  John  B.  (M  1923),  Pres.  and  Treas.  (for 

mail),   John   B.   Feehan,    Inc.,   471   Union  St., 

Lynn,  and  4  Ocean  View  Dr.,  Marblehead,  Mass, 
FEELY,   Frank  J.   (.4   1929),   17215  Greenlawn, 

Detroit,  Mich, 
FEGLEY,   Donald  R.    (/   1935;   5   1933),   2274 

Loring  PL,  New  York,  N.  Y. 
FEHUG,  John  B.  (M  1918),  Pres.  (for  mail), 

Excelsior  Htg.  Supply  Co.,  528  Delaware  St.,  and 

2927  Brooklyn  Ave.,  Kansas  City,  Mo. 
FELDMAN,    A.    M.*    (Life    Member;    M    1903), 

Consulting  Engr.,  40  West  77th  St.,  New  York, 

N.  Y 
FELS,  Arthur  B.   (M  1919),  The  Fels  Co.,  42 

Union  St.,  Portland,  Maine. 
FELTWELL,  Robert  H.  (M  1005),  Htg,  Engr., 

U.  S.  Radiator  Corp.,  2321  Fourth  St.  N.E.,  and 

(for  mail),  1370  Oak  St.  N.W.,  Washington,  D.  C. 
FENNER,  N.  Paul  (A  192&),  Hoffman  Specialty 

Co.,  500  Fifth  Ave.,  Room  3324,  New  York,  N.  Y. 
FENSTERMAKER,  Sidney  E.  (M  1909),  Pres. 

(for   mail),    S.    E.    Fenstermaker   &    Co.,    937 

Architects  and  Builders  Bids;.,  and  3102  Washing- 

ton Blvd.,  Indianapolis,  Ind. 
FERGUSON,  Ralph  R.  (M  1934;  A  1927;  J  1925). 

Mgr.  Trade  Dept.,  American  Blower  Corp.,  401 

Broadway,  New  York,  N.  Y.,  and  (for  mail), 

100  Prospect  St.,  Kast  Orange,  N.  J. 
FERNALD,  Henry  B.,  Jr.  (J  193f>:  S  1933),  145 

Lorraine  Ave.,  Upper  Moutelair,  N.  J. 
FERRERO,  Henry  J.   (J   1935;  6'  1933),   1738 

Adams  St.,  New  York,  N.  Y. 
FIEDLER,  karry  William  (M  1923),  Pres.  (for 

mail),  Air  Conditioning  Utilities,  Inc.,  480  Fifth 

Ave.,  New  York,  and  49  Palmer  Ave.»  Scaradale, 

N.Y, 
FIFE,  George  I).  (A  1931;  J  1929),  Kngr.,  Air 

Cond.,  National  Broadcasting  Co.,  30  Rocke- 

feller Plaza,  and  (for  mail),  102  Kast  2Snd  St., 

New  York,  N.  Y. 
FILKINS,  Harry  L.  (A  1933),  Vice-Pres.,  City  Ice 

Co.  of  Kansas  City,  21st  and  Campbell  Sta.,  and 

(for  mail),  34  Kast  55  Terrace,  Kansas  City,  Mo. 
FltLO,  Frank  B.  (A  -1034),  Minneapolis-  Honey- 

well Regulator  Co.,  2831  Olive  St.,  St,  Louis,  Mo. 
FIN  AN,  James  Jf.  (H  1923),  Supervising  Engr., 

Board  of  Education,  City  of  Chicago,  228  N. 

LaSalle   St.,    Builders   Bldg,,   and    (for   mail), 

7149  Euclid  Ave.,  Chicago,  III, 
FINCH,  Stanley  k  (A  1931),  Industrial  Engr., 

Brooklyn   Union   Caa   Co.,    180   Kemsen   $t.» 

Brooklyn,  N.  Y. 
FIRESTONE,  Jamoa  F.  (A  1025;  J  1914),  Exec. 

Vice-Pros.*  Round  Oak  Furnace  Co,,  and  (for 

mail),  203  Orchard  St.»  Dowagl&e*  Mich, 
FITTS,  Ghnrlea  JD.  (M  ItKZOl,  Mgr.  (for  mail), 

American  Radiator  Co.,  692  Prior  Ave.t  St.  Paul, 

and  2807  Dean  Blvd.,  Minneapolis,  Mirm,- 
HTTS,  Joseph  €.   (M   1930),  Secy..   Heating, 

Piping    ana     Air    Conditioning     Contractor! 

National  Assn.,  1260  Sixth  Ave,»  New  York* 

N.  Y.,  and  (for  mail),  215  Kenilworth  Rd,» 


Mat 


atthew  J.  (M  1034),  Tre 
Standard  Asbestos  Mfg.  Con  820  W,  Lake  St» 
and  (for  mail),  7314  Hanwd  Ave,»  CMctgQ,  I1L 

FITZSIMONS,  J.  Patrick  C/  19S4;  S  1982),  The 
Robert  FiUalmons  Co,»  Ltd.,  21  Rebecca  St*« 
Hamilton,  Ont.(  Canada, 

FLANAGAN,  Edward  T*  (A  1929),  C.  A.  Dunham 
Co.,  Ltd.,  1139  Bay  St.,  Toronto,  Ont.»  Caimda. 

FLARSHEIM,  Clnr*mc«  A.  (J  1933),  Mgr.>  ^ir 
Cond.  DepU  Stewart  Warner-Atemite  '  Co., 
2425  McGee  St*»  and  (for  mail),  8720  Holmes 
St«»  Kansas  City,  Mo. 

FLBISHBH,  Walter  JU*  <M  1014),  Consulting 
Engr.  (for  mall),  11  West  42nd  &t,»  New  Yorfc, 
and  Saw  Mill  F&rm,  New  City,  N.  Y. 


16 


ROLL  OF  MEMBERSHIP 


FLEMING,  James  P.  (M  1923),  Engr.- Custodian, 
Board  of  Education,  5045  N.  Kimball  Ave., 
Chicago,  111. 

FLINK,  Carl  H.  (M  1923),  Director  of  Research 
(for  mail),  American  Gas  Products  Corp.,  408 
East  lllth  St.,  New  York,  and  74  Brookside 
Ave.,  Mt.  Vernon,  N.  Y. 

FLINT,  Coll  T.  (M  1919),  N.E.  Sales  Mgr.  (for 
mail),  The  H.  B.  Smith  Co.,  640  Main  St., 
Cambridge,  and  56  Brantwood  Rd.,  Arlington, 
Mass. 

FLOYD,  Morris  (M  1933),  Mgr.,  Air  Cond.  Div., 
Edwards  Mfg.  Co.,  Cincinnati,  Ohio. 

FOGARTY,  Orville  A.  (M  1934),  Mgr.,  Oil 
Burner  Div.,  Canadian  Fairbanks-Morse  Co., 
Ltd.,  980  St.  Antonie  St.,  and  (for  mail),  2178 
Old  Orchard  Ave.,  Montreal,  Que.,  Canada. 

FONDA,  Bayard  P.  (M  1934),  Air  Cond.  Engr. 
(for  mail),  Bryant  Heater  Co.,  17825  St.  Clair 
Ave.,  Cleveland,  and  2905  Hampton  Rd., 
Shaker  Heights,  Ohio. 

FORFAR,  Donald  M.  (M  1917),  Mech.  Engr.  (for 
mail),  Grinnell  Co.,  240  Seventh  Ave.  S.,  and 
4817  Emerson  Ave.  S.,  Minneapolis,  Minn. 

FORSBERG,  William  (M  1919),  Hopson  & 
Chapin  Mfg.  Co.,  231  State  St.,  New  London, 

FORSYTH,  Arthur  T.  (A  1934),  Dist.  Repr., 
Buffalo  Forge  Co.,  2434  First  Ave.  S.,  Seattle, 
Wash. 

FOSTER,  Charles  (M  1923),  Consulting  Engr. 
(for  mail),  508  Sellwood  Bldg.,  and  2831  E. 
First  St.,  Duluth,  Minn. 

FOSTER,  James  M.  (M  1930;  A  1920),  Factory 
Repr.  (for  mail),  4526  Olive  St.,  St.  Louis,  and 
7021  Lindell  Ave.,  University  City,  Mo. 

FOSTER,  Tillman  R.  (J  1930),  Carrier  Engrg. 
Corp.,  180  N.  Michigan  Ave.>  Chicago,  111, 

FOUL0S,  P.  A.  L.  (M  1915),  Mech.  Engr.  (for 
mail),  Office  of  Holhs  French,  Consulting  Engr., 
210  South  St.,  Boston,  and  72  Whitin  Ave., 
Point  of  Pines,  Revere,  Mass. 

FOULDS,  Samuel  T.  N.  (J  1930),  Sales  Engr., 
Power  Equipment  Co.,  791  Trcmont  St.,  Boston, 
and  (for  mail),  72  Whitin  Ave.,  Revere,  Mass. 

FOWLES,  Harry  H.  (J  1934),  Heating  Engr., 
Carman-Thompson  Co.,  12-14  Lincoln  St., 
Lewiston,  and  (for  mail),  Y.  M.  C.  A.,  Auburn, 
Maine*. 

FOX,  Otto  (M  1931),  Chief  Engr.  (for  mail), 
Bryant  Heater  Co.,  17825  St.  Clair  Ave.,  Cleve- 
land, and  1819  Fannington  Rd.,  East  Cleveland, 
Ohio. 

FRAMPTON,  Alfred  C»  (S  1984),  729>lJ  Wilson, 
Norman,  Okla. 

FRANK,  John  M.  (M  1918:  A  1912),  Il«  Elec. 
Vtg,  Co..  2850  N.  Crawford  Ave.,  Chicago,  111. 

FRANK,  Olive  B.*  (Af  1010),  Pwa.  (for  mail), 
Frank  Bngrg.  Co.,  U  Park  PI.,  and  600  West 
114th  St.,  New  York,  N.  Y. 

FRANKEL,  Gilbert  S.  (M  1920).  Mgr.,  Federal 
&  Marine  Dcpt.  (for  mail),  Buffalo  Forge  Co., 
403  Commercial  National  Bank  Bldg.,  and  2749 
Macomb  St.  N.W.,  Washington,  D,  C, 

FRANKLIN,  Ralph  S.  (M  1019),  Prea*Treas.  (for 
mail),  Albert  B.  KrankUn,  Inc,»  38  Chauncy  St., 
Boston*  and  320  Grove  St.,  Melrose,  Mass. 

FREAS,  Royal  Bruce  (M  1928),  Pres.  (for  mall), 
Freas  Thermo  Electric  Co.,  1206  S,  Grove  St., 
Irvington,  N.  J.,  and  4  West  43rd  St.,  New  York, 
N.  Y. 

FREEMAN,  Alton*  M.  (A  1©29)>  Sales  Engr., 
6088  Plankittgton  Bidg»,  and  (for  mail),  4533  N. 
Bftrtlett  Ave.,  Milwaukee,  Wla, 

FREITAG,  Frederic  G,  (M  1032),  Consulting 
Hngr,  (for  mail),  Sylvestre  Oil  Co.,  709  S. 
Columbus  Ave^  and  9  Harrison  St.,  Mt.  Vernon, 
N.  Y, 

FRENCH,  Donald  (M  1926),  Vice-Pres.  (for 
mail).  Carrier  Corp,,  850  Frelinffhuysen  Ave., 
Newark,  and  40  Waldron  Ave,,  Summit,  N.  J. 

FRBY,  Georftti  O.  (J  1984),  Elec.  Bnjpr.  (for  mail), 
Warner  Brou.  Theatres,  Inc.,  mi  West  44th  St., 
New  York,  and  221  Linden  Blvd.,  Brooklyn, 
N.Y, 


FRIEDMAN,  Ferdinand  J.  (M  1921),  McDougall 

&  Friedman,  31  Union  Square,  New  York,  N.  Y.,  , 

and    (for   mail),    1221   Osborne   St.,    Montreal, 

Que.,  Canada. 
FRIEDMAN,  Milton   (S  1933),  470  West  End 

Ave.,  New  York,  N.  Y. 
FRITZ,   Charles   V.    (S  1933),   P.   O.    Box  303, 

Carnegie   Institute  of  Technology,   Pittsburgh, 

Pa. 
FRITZBERG,  L.  Hilding  (/  1931),  Engr.   (for 

mail),  B.  F.  Sturtevant  Co.,  and  1338  River  St., 

Hyde  Park,  Boston,  Mass. 
FUKUI,    Kunitaro   (M   1920),   Oriental   Carrier 

Engrg.  Co.,  Ltd.,  Osaka  Mitsui  Bldg.,  Nakano- 

shima,  Osaka,  Japan. 


GABY,  Frederick  A.  (M  1926),  Chief  Engr.  (for 
mail),  Hydro-Electric  Power  Commission  of 
Ontario,  190  University  Ave.,  and  480  Spadina 
Rd.,  Toronto,  Ont.,  Canada. 

GALLIGAN,  Andrew  B.  (M  1921),  716  South 
51st  St.,  Philadelphia,  Pa. 

GALLOWAY,  James  F.  (S  1934),  176  Clarkson 
Ave.,  Brooklyn,  N.  Y. 

GAMMILL,  Oscar  E.,  Jr.  (J  1930),  Sales  Engr. 
(for  mail),  Carrier  Engrg.  Corp.,  1416  Hibernia 
Bank  Bldg.,  and  2133  Calhoun  St.,  New  Orleans, 
La. 

GANT,  H.  P.*  (M  1915),  (Presidential  Member}, 
(Pres.,  1923;  1st  Vice-Pres.,  1922;  2nd  Vice- 
Pres.,  1921;  Council,  1918-1924),  Vice-Pres.  (for 
mail),  Carrier  Engrg.  Corp.,  12  South  12th  St., 
and  Penn  Athletic  Club,  Philadelphia,  Pa. 

GARDNER,  S.  Franklin  (M  1911),  Pres.  (for 
mail),  Standard  Engrg.  Co.,  2129  Eye  St.  N.W., 
and  4901  Hillbrook  Lane,  Washington,  D.  C. 

GARDNER,  William,  Jr.  (A  1921),  Vice-Pres. 
(for  mail),  Garden  City  Fan  Co.,  1842  McCor- 
mick  Bldg.,  and  7836  Loomis  Blvd.,  Chicago,  111. 

GARNEAU,  L6o  (J  1930),  Sales  Engr.,  Room  743 
Dominion  Square  Bldg.,  and  (for  mail),  8454 
Brouages  St.,  Montreal,  P.  Q.,  Canada, 

GAULT,  George  W.  (S  1934),  Marysville,  Pa. 

GAUSMAN,  Carl  E.  (M  1923),  Mech.  Engr., 
1100  Minnesota  Bldg.,  and  (for  mail),  2360 
Chilcombe  Ave.,  St.  Paul,  Minn. 

GAUTESEN,  AH  (J  1935;  5  1933),  1039-79th  St., 
Brooklyn,  N.  Y. 

GAWTHROP,  Fred  H.  (M  1919),  Pres.,  Gawthrop 
&  Bro.  Co.,  705  Orange  St.,  and  (for  mail),  2211 
Shallcross  Ave.,  Wilmington,  Del. 

GAY,  Lewis  M.  (A  1934),  Power  Engr.  (for  mail), 
Texas  Power  &  Light  Co.,  Box  902,  Dallas,  and 
724  Griffith  Ave.,  Terrell,  Texas, 

GAYLOR,  William  S.  (M  1919),  Consulting 
Kngr,,  Flameklng  Co.,  Inc.,  2159  Madison  Ave,, 
New  York,  and  (for  mail),  42  Mayhew  Ave., 
Larchmont,  N.  Y. 

GAYLORD,  F.  H.  (M  1921),  Western  Sales  Mgr. 
(for  mail),  Hoffman  Specialty  Co.,  Inc.,  130  N. 
Wells  St.,  Chicago,  and  362  N.  York  St.,  Elm- 
hurst,  111. 

GEIGER,  Irvin  H.  (M  1919),  Reg.  Prof.  Engr.  and 
Mfrs.  Repr,,  Room  319  Telegraph  Bldg.,  Har- 
risburg,  Pa, 

GEISSBUHLER,  John  O,  (S  1034),  University 
Circle,  and  (for  mall),  9820  Zirnmer  Ave., 
Cleveland,  Ohio. 

GELB,  Amiel  CS  1935),  1042  Irving  Ave.  N,, 
Minneapolis,  Minn, 

GENCHI,  Bernard  (J  1935;  5  1938),  8808~15th 
Ave.,  Brooklyn,  N.  Y. 

GERMAIN,  Oscar  (M  1935),  Germain  Frere, 
Ltd.,  1343  Blvd.  St.  Louis,  Three  Rivers,  Que., 
Canada. 

GERRISH,  Grenvllle  B.  (/  1930),  Mgr..  Fitz- 
gibbona  Boiler  Co.,  Inc. ,,80  Boylston  St.,  Boston, 
and  (for  mail),  1  Overlook  Rd.,  Melroee,  Mass. 

GERRISH,  Harry  E.  (M  1610),  (Coundl,  1919), 
Vice-Frea,  (for  mail),  Morgan-Gerrish  Co.,  807 
Essex  Bldg,,  and  4584  Fremont  St.,  Minneapolis, 
Minn. 


17- 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


GESMER,    Joseph    (J    1935;    5    1933),    Student 

Engr.,  B.  F.  Sturtevant  Co.,  Hyde  Park,  and  (for 

mail),  41  Beacon  St.,  Quincy,  Mass. 
GETSGHOW,   George  M.   (Ad    1906),  Pres.  and 

Treas.  (for  mail),  Phillips  Getschow  Co.,  32  W. 

Austin  Ave.,  and  4542  Beacon  St.,  Chicago,  111. 
GETSCHOW,  Roy  M.  (M  1919),  Secy,  (for  mail), 

Phillips  Getschow  Co.,  32  W.  Austin  Ave.,  and 

1336  Arthur  Ave.,  Chicago,  111. 
GIANNINI,  Albert  A.  (/  1935;  5  1933),  64  West 

170th  St.,  New  York,  N.  Y. 
GIBBS,  Edward  W.  (At  1919),  (for  mail),  Smith- 

Gibbs  Co.,  201  S.  Alain  St.,  and  234  President 

Ave.,  Providence,  R.  I. 
GIBBS,  Frank  C.  (M  1921),  Gen.  Supt.  (for  mail), 

National     Regulator     Co.,     2301     ICnox    Ave., 

Chicago,  and  150  N.  Cuyler  Ave.,  Oak  Park,  111. 
GIESEGKE,  Frederick  E.*  (M  1913),  (Council, 

1932-1934),  Director,  Texas  Engrg.  Experiment 

Station,  Agricultural  and  Mechanical  College  of 

Texas,  College  Station,  Texas. 
GIFFORD,    Clarence    A.    (A    1934),    Salesman, 

American  Radiator  Co.,  374  Delaware  Ave,,  and 

(for  mail),  247  North  Dr.,  Buffalo,  N.  Y. 
GIFFORD,  Robert  L,  (M  1908),  Pres.,  Illinois 

Engrg.  Co.,  21st  St.  and  Racine  Ave.,  Chicago, 

III.,  and  (for  mail),   1231  S.  El  Molino  Ave., 

Pasadena,  Calif. 
GIGUERE,  George  H.  (M  1920),  Mech.  Engr., 

800   Marquette    Bldg.,   and    (for   mail),    17205 

Fairport  Ave.,  Detroit,  Mich. 
GILES,  Alfred  F.  (J  1934),  H.  H.  Robertson  Co., 

2000  Grant  Bldg.,  and  (for  mail),  4307  Ludwick 

St.,  Pittsburgh,  Pa. 
GILFRIN,  George  F.  (M  1932),  Gen.  Repr.  (for 

mail),     Carrier- Brunswick-International,     Inc., 

Apartado  03,  Bis  and  Paseo  tie  la  Ret'orma  120, 

Lomas  de  Chapultepec,  Mexico,  D,  F. 
GILL,   John   W.    (S    1935),   2120   Carter   Ave., 

St,  Paul,  Minn. 
GILLE,  Hadar  B.  (A/  1930),  Skoldungagatan  4, 

Stockholm,  Sweden. 
GILLETT,  M.  C.  (M  191(5),  Engr.,  OUOO  Rising 

Sun  Ave.,  Philadelphia,  Pu. 
G1LLHAM,  Walter  K.  (/U  1917),  (Trcas.,  1920- 

1929;  Council,  19120-19129),  Consulting  Engr,  (for 

mail),  814  Inter-State  Bids,,  and  3427  Belief on- 

taine  Ave.,  Kansas  City,  Mo. 
GILLING,  William  F.,  Jr.  (Life  Member;  M  1933; 

A   1919),  Aast.  Mgr.,  American  Kadiator  Co., 

127   Federal   St.,    Boston,  and    (for   mail),   2tt 

Abbott  Rd.,  Wellesley  Hills,  Mass. 
GILMAN,  Franklin  W.  (M  1935),  Plant  Engr, 

(for  mail),  Atwater  Kent  Mfg.  Co,,  4700  wfa- 

aahickon  Ave.,  and  5H  W,  Coulter  St.,  Phila- 
delphia, Pa, 
GILMQRE,  Louis  A.  (J  1935;  5  1930),  Vice-Pres. 

(for  mail),  John  Gilmore  &  Co.,  13  North  10th 

St.,  and  01SG  Westminster  PI.,  St.  Louis,  Mo, 
GILMOUR,  Alan  B.  (A  1032),  Salesman,  B.  K. 

Gilmour  Co.,  Inc.,  15B-41at  St.,  and  (for  mail), 

410  Ocean  Ave,,  Brooklyn,  N.  Y. 
GINJ,  Aido  (M  1933),  via  Correggio  18,  Milano, 

Italy. 
GIVIN,  Albert  W.  (,-1  ittiifl),  The  Guraey  Foundry 

Co,,  Ltd,,  P,  O,  Box  1149,  Montreal,  P.  &„ 

Canada, 

GLANZ,  Edward  (A  1930),  Prea.  (for  mail), 
Glanz  &  Killian  Co,,  1701  W.  Forrest  Ave,>  and 
3865  Lake  wood  Ave.,  Detroit,  Mich, 
GLASS,  William  (A/  1934),  Mgr,  (for  mall). 
Partridge- Halli day,  Ltd.,  144  Lombard  St,, 
Winnipeg,  and  190  Braemar  Ave,,  Norwood, 
Manitoba,  Canada, 

GLASSKY,  J,  Wilbur  (M  1922),  Partner  (for 
mail),  Vapor  Kngrg.  Co,,  10  South  18th  St, 
Philadelphia,  and  7818  Ardleigh  St.,  Chestnut 
Hill,  Philadelphia,  Pa. 

GLEASON,  Gilbert  H,  (M  1023),  Partner  (for 
mail),  Gilbert  Howe  Gieason  &  Co,,  2&  Hunt- 
ington  Ave,  Boston,  and  10  Edgehlll  Kd,» 
Winchester,  Mass, 

GtOBJL  Evlns  Foree  (A  1916).  Pres.,  Evins  F, 
Glore  Sales  Corp.,  1949  Grand  Central  Terrnin&U 
and  (for  mail),  644  Riverside  Dr.,  New  York, 
N.  Y, 


GOELZ,  Arnold  H.  (M  1U31),  Pres.  (for  mail), 
ICrposchell  Engrg.  Co.,  2300  N.  Knox  Ave., 
Chicago,  and  8U7  Greenwood  Ave.,  Wilmettc,  111. 

GOENAGA,  Ro&er  C.  (M  1031),  Tech.  Director 
(for  mail),  Ateliers  Ventil.,  100  Cours  Gambctta, 
Lyon,  and  33  Avenue  Valioud,  Ste  Foy-les-lyon, 
Rhone,  France. 

GQERG,  B.  (M  1928),  (for  mail),  American 
Radiator  Co.,  075  Bronx  River  Rd.,  Yonkers, 
and  294  Bionxville  Rd.,  Bronxville,  N.  Y. 

GOLDBERG,  Moses  (.-1  1934),  Pres.,  Electric 
Motors  Corp.,  108  Centre  St.,  New  York,  and 
(for  mail),  13 U  E.  Seventh  St.,  Brooklyn,  N.  Y. 

GOLDSGHM1DT,  Otto  E.  (M  1«1.">),  Consulting 
Engr.  (for  mail),  110  West  10th  St.,  and  345 
Kast  57th  St.,  New  York,  N.  Y. 


Ave.,  New  York  City  Assn.,  1045  Grand  Central 

Terminal  Bldg.,  New  York,  N.  Y.,  and  (for  mail), 

1(50  Halsted  St.,  Kast  Orange  N,  ]. 
GOODRICH,  Charles  F.  (A/  1010),  Andrews  & 

Goodrich,    Inc.,    Boston,    and    (for    mail),    330 

Adams  St.,  Dorchester,  Mass. 
GOODWIN,   Samuel  L.    (4U   1024),   Consulting 

Engr,,  247  Madison  Ave.,  Ilasbrouck  Heights, 

N.  J. 
3DV 


Pittsburgh,  and  0032  Marie  St.,  Pittsburgh,  Pa. 
GORDON,    Edward   B.,    Jr.    (M    1008),    Pres., 

Pillsbury  Engrg.  Co.,  1200  Second  Ave,,  and  (for 

mail),  2450  West  24th  St.,  Minneapolis,  Minn, 
GORDON,  Peter  B.  (J  1035),  Kngr.  (for  mail), 

George  E.  Gibson  Co.,  441  Lexington  Ave,,  New 

York,  N,  Y.,  and  35  Park  Ave.,  Bloomfidd,  N.  T. 
GORDON,  William  J,,  4r,  (6'  1035),  2208  Oliver 

Ave,  S.,  Minneapolis,  Minn, 
GORNSTON,  Michael  II.  (A  1023),  Stationary 

Engr,  (for  mail),  430  Dumont  Ave.,  Brooklyn, 

and  8504  Woodhaven  Blvd.,  Woodhaven,  N.  Y. 
GOSSETT,  Earl  J.  (M  1023),  Pres,  (for  mail), 

Bell  &  Goaaett  Co.,  3000  Wallace  St.,  Chicago, 

and  314  Woodland  Ave.,  Winnetka,  III, 
GOTTWALD,  C.  (A  1910).  Pres,  (for  mail),  The 

RiCMvIL  Co.,  Union  Trust  Bldg,,  Cleveland,  and 

2225  Stillman  Rd.,  Cleveland  Heights,  Ohio. 
GOULDING,  William  (A   1933),  Kn«r«.   Dent., 

National   Broadcasting  Co.,   Radio   City.   New 

York,  and  (for  mail),  409  Kast  17th  St,,  Brook- 

lyn, N.  Y. 
GRAHAM,  Charles  H.  (M  1034),  Sales  Engr,, 

Lennox  Furnace  Co.,  Inc.,  Syracuse,  and   (lor 

mail),  93  Lake  St.,  Hamburg,  N,  Y, 
GRAHAM,  William  D.  (M  1920:  A  1925;  /  1923), 

Dist.  Mgr,  (for  mail),  Carrier  Engrg.  Corp,,  8t>0 

Union  Trust  Bldg,,  Cleveland,  Ohio, 
GRAHN,  Victor  F,   (M  1927),   Htg.  and   Vtg, 

Engr.,  fcnney  &  Ohmea,  Inc.,  101  Park  Ave., 

New  York,  N,  Y.,  and  (for  mail),  120  Greenwood 

Ave,,  East  Orange,  N,  J. 
GRA.NSTON,  Ray  O.  (/  1935;  S  1030),  Bn«r.> 

Univ.  Plbfr  &  Htg.  Co,,  3039  University  Way, 

and  (for  mail),  4568  Fourth  Ave,  N,E,,  Seattle, 

Wash. 
GRAJNT,  Walter  A.  (A  1933;  /  1929),  Develop- 

ment   Engr.,   Carrier  Eagrg.   Corp,,    750  Fre- 

Unghuysen  Ave,,  Newark,  and  (for  mnil),  H&Q 

Anna  St,,  Elizabeth,  N,  J. 
GRAVES,  WUiurd  B,  (Life  Mmfar;  M  1000), 

Pr«d,  (for  mail),  W,  B,  Graves  Htg,  Co,,  102  N 
,,  Chicago,  III 


GRAY,  E&rte  W.  (A  1834),  Comm©rdai  Otpt.,  la 
Charge  of  Air  Cond,  Sales  (for  mail).  Oklahoma 
Gaa  &  Elec,  Co,,  Box  1408,  aad  2125  N.W., 
18th,  Oklahoma  City,  Qkla* 

GRAY,  Gyorfle  A.  (M  l«M)f  C,  A,  Duttham  Co,, 
Ltd,,  404  Flam  BWg,»  Ottawa,  OntM  Canada, 

GRAY,   Wlliiam   E,    (M   IMS),    Saiei 
Powers  Regulator  Co.,  2^720  Gr0^ftvl«w 
j  m.,  and  (for  mail),  Bw  2 


18 


ROLL  OF  MEMBERSHIP 


GREEN,  Joseph  J.  (A  1938),  Mfrs.  Repr.,  Joseph 

J.  Green  Co.,   Buffalo,  and  (for  mail),  328  W. 

Girard  Blvd.,  Kenmore,  N.  Y. 
GREEN,   William  C.    (Life   Member;    M   1906), 

Dist.  Mgr.   (for  mail),  Warren  Webster  &  Co., 

704    Race    St.,    and    244    Erkenbrecher    Ave. 

(Avondale),  Cincinnati,  Ohio. 
GREENBURG,  Dr.  Leonard"  (M  1932),  Acting 

Health  Officer  (for  mail),  New  Haven  Dept.  of 

Health,    City  Hall,    161   Church  St.,   and   519 

George  St.,  New  Haven,  Conn. 
GREENLAND,  Sidney  F.   (M  1034),   Htg.  and 

Vtg.  Engr.,  Gee,  Walker  &  Slater,  Ltd.,  32  St. 

James  St.,  London,  S.W.  1,  and  (for  mail),  71 

Arodene  Rd.,  Bnxton,  London  S.W.  2,  England. 
GREER,   Willis  R.    (J   1934),  Air  Cond.   Engr., 

Arkansas  Power  &  Light  Co.,  and   (for  mail), 
'  1401  Linden  St.,  Pine  Bluff,  Ark. 
GRIFFIN,  DeWitt  C.  (M  1933),  Secy-Treas.  (for 

mail),  May  &  Griffin,  Inc.,  501  Orpheum  Bldg., 

and  9717~47th  S.W.,  Seattle,  Wash. 
GRIFFIN,  John  ,T.  (M  1921;  A  11)18),  Vice-Pres. 

and  Dir.  (for  mail),  Mutual  Bank  and  Trust  Co., 

710  Locust  St.,  and  3S52  Castlcman  Ave.,  St. 

Louis,  Mo. 
GROSECLOSE,     John     B.     (A     1929),     Engr., 

Estimator,  Dixie  Htg,  &  Vtg.  Co.,  109  Fannin 

St.,    and    (for    mail),    3424    University    Blvd., 

Houston,  Texas. 
GROSS,  Lyman  C.  (M  1931),  Consulting  Engr. 

4G53-13th  Ave.  S.,  Minneapolis,  Minn. 
GROSSMAN,  Harry  E.  (.4  1033;  J  1927),  Sales 

Repr.,  Haynea  Selling  Co.,  Inc.,  1518  Fairmount 

Ave.,  Philadelphia,  and  (for  mail),  405  Custer 

Ave.,  Glenoldcn,  Pa, 
GROSSMANN,    Harry    A.    (M    1931),    H,    A. 

Groaamann  Co.,  3221  Olive  St.,  and  (for  mail), 

3122  Gcyer  Ave.,  St.  Louis,  Mo. 
GUNTIIER,  Felix  A.*  (M  1925),  Sales  Kngr.  (for 

mail),  429- B  Oliver  Bldg.,  and  Box  220  R.  D.  0, 

S.  Hills  Branch,  Pittsburgh,  Pa. 
GURNKY,    Edward    Holt    (M    1020),    (Council, 

1031-15)34),  Pres,  (for  mail),  Gurney  Foundry 

Co.,  Ltd.,  4  Junction  Rd.,  and  347  Walmer  Rd., 

Toronto,  Ont,  Canada. 

H 

HAAS,  Emii,  Jr.  (J  1920),  Secy-Treas.  (for  mail), 

Natkin  &  Co,,  2020  Wyandotte,  and  Ncwbern 

Hotel,  Kansas  City,  Mo, 

HAAS,  Samuel  L.  (M  1023),  Pros,   (for  mail), 
-  Advance  Heating  Co.,  117-1U  N.  DespUuncs  St., 

and  1T>13  Furtfo  Ave.,  Chicago,  111. 
HAATVEDT,   Sheldon   R.    (,V    1035),   ai5-10th 

Ave.  S.IC.,  Minneapolis,  Minn. 
HACKKTT,  H.  Berkeley  at  1021),  901  Architects 

BUlg..  17th  and  Sansom  Stsu,  Philadelphia,  Pa. 
HADDOCK,  Isaac  IV (A  1026),  New  England  Gas 

&  Elec.  Assn.,  710  Maaaachusetty  Ave.,  Cam- 
bridge, Masi. 
HADEN,  G.  Nelacm  (M  1034;  A  1928;  /  1922), 

Director  (for  mail),  G.  N.  Haden  &  vSona,  Ltd., 

Lincoln  House,  00  Kingeway,  London  W,  C.  2, 

and    36    Wild  wood    Rd.,    Hampsteart    Heath, 

London  N.W,  1L  England, 
HADEN,  William  Nelson  (Lift  Mtmbv;  M  1902), 

Late  Chairman,  G.  N,  Haden  &  Sons,  Ltd.,  St 

Georges  Works,  and  (for  mail),  Arnolds  Hill, 

Trowbrldge,  Wilt,  England, 
HADESTY,  Alfred  L.,  Jr.   (M  19*1),  130  E. 

Broad  St,t  Tamnqua,  Pa, 
HADJISKY,  Joseph  N,  <M  1980),  Consulting 

Engr.,  744  Bates  St.,  Birmingham,  Mich, 
HAGAN,  WliUam  V.  (A  1933;  J  3026),  Secy,  (for 

mail),  608  Pearl  St.,  and  1811  Jones  St.,  Sioux 

City,  Iowa. 
HAGfDON*  Charles  H*  (M  1919),  S,  E,  Fenatcf* 

maker  &  Go,,  W  Architect!  &  Builder®  Bldg., 

indiawtpoli*,  tod. 
HAIGNEV,  Jfolm  B,  (J  193#t  S  1988),  8621  Shore 

Ed,,  Brooklyn,  N,,Y, 
HAfNftSI*  Jobia  J*  (M  1915).  Free,  (lor  mail),  Th« 

HaineT  Co,,  19*5  W,  Late  St.,  Chicago,  and 

S$0~J7th  Av«,»  Maywood,  111, 


HAJEK,  William  J.  (M  1932),  Br.  Mgr.  (for  mail), 
Minneapolis-Honeywell  Regulator  Co.,  285  Co- 
lumbus Ave.,  and  333  Beacon  St.,  Boston,  Mass. 
HAKES,  Leon  M.  (M  1932;  A  1932;  J  1929),  Sales 
Engr.  (for  mail),  The  R.  T.  Coe  Co.,  400  Rey- 
nolds Arcade  Bldg.,  and  71  Stratmore  Dr.- 
Greece,  Rochester,  N.  Y. 

HALE,  John  F.  (M  1902),  (Presidential  Member), 
(Pres.,  1913;  1st  Vice-Pres.,  1912;  Board  of 
Governors,  1908-1910,  1912-1913),  Dist.  Mgr. 
(for  mail),  Aerofin  Corp.,  Ill  W.  Washington  St., 
Rm.  1058,  Chicago,  and  408  S.  Brainard  Ave., 
LaGrange,  111. 

HALEY,  Harry  S.*  (M  1914),  Consulting  Engr., 
Partner  (for  mail),  Leland  &  Haley,  58  Sutter  St., 
and  735-21st  Ave.,  San  Francisco,  Calif, 

HALL,  John  R.  (J  1032),  Mech.  Engr.,  U.  S.  Air 
Cond,  Corp.,  2101  N.E.  Kennedy  St.,  and  (for 
mail),  141U  Lakeview  Ave.,  Minneapolis,  Minn. 

HALL,  Mora  S.  (M  1934),  Combustion  Engr,  (for 
mail),  May  Oil  Burner  Corp.,  Maryland  and 
Oliver  St.,  Baltimore,  and  Route  No.  3,  West- 
minster, Md. 

HAMBURGER,  Fred  G.  (J  1935;  5  1933),  185 
West  102nd  St.,  New  York,  N.  Y. 

HAMENT,  Louis  (A  1933),  Mgr.  (for  mail),  Aqu- 
atic Chemical  &  Metallurgical  Engrs.,  118  East 
28th  St.,  and  568  East  166th  St.,  New  York.  N.Y. 

HAMERSKI,  Francis  D.  (J  1934),  626  E.  Fifth 
St.,  Winona,  Minn. 

HAMILTON,  James  E.  (A  1933),  Mgr.  (for  mail), 
U.  S.  Radiator  Corp.,  4004  Duncan  Ave.,  St. 
Louis,  and  7715  Shirley  Dr.,  Clayton,  Mo. 

HAMUN,  Chauncey  J.T  Jr.  (.4  1934),  Harnlin 
Air  Conditioning  Co.,  and  (for  mail),  1014 
Delaware  Ave.,  Buffalo,  N.  Y. 

HAMLIN,  Harry  A.  (A  1910),  Br.  Mgr.  (for  mail), 
Johnson  Service  Co.,  427  Brainard  St.,  Detroit, 
and  120  Winona,  Highland  Park,  Mich. 

HANLEY,  Edward  V.  (A  1933),  Pres.  (for  mail), 
S.  V.  Hanley  Co.,  1653  N.  Farwell  Ave.,  Milwau- 
kee, and  844  E.  Birch  Ave.,  Whitefish  Bay,  Wis. 

HANLEY,  Thomas  F.,  Jr,  (M  1933),  Pres.  (for 
mail),  Hanley  &  Co.,  1503  S.  Michigan  Ave.,  and 
4940  East  End  Ave.,  Chicago,  III. 

HANSEN,  Carl  J.  (J  1935;  S  1933),  273  Sheffield 
Rd.,  Lansclowne,  Pa. 

HANSEN,  Charles  C.  (M  1928),  Engr.,  428 
Prospect  Sta.,  South  Orange,  N.  J. 

HANSON,  Leslie  P.  (J  1935;  S  1033),  Engr.,  U.  S. 
Air  Cond.  Corp.,  and  (for  mail),  43«JB-46th  Ave, 
S.,  Minneapolis,  Minn. 

HARDING,  Louis  A,*  (M  1911),  (Presidential 
Member'),  (Pres.,  1930;  1st  Vice-Pres.,  1929; 
2nd  Vice- Pres.,  1928;  Council,  1922-1931),  Prea, 
(for  mail),  L.  A,  Harding  Construction  Corp., 
Prudential  Bldg.,  and  85  Cleveland  Ave.,  Buffalo, 
N.  V. 

HARE,  W.  Almon  (M  1930),  Pres.,  Hare  Stoker 
Corp..  4853  Rivard  St.,  Detroit,  Mich. 

HARMS,  William  T.*  (M  1917),  1015  Vlnewood 
Ave.,  Detroit,  Mich. 

HARRIGAN,  Edward  M.  (M  1015),  (for  mail), 
Harrigan  &  Keid  Co,,  1366  Bagley  Ave.,  and 
7460  LaSalle  Blvd.,  Detroit,  Mich. 

HARRINGTON,  Charles  (M  1923),  43  Indian 
Grove,  Toronto,  Ont,  Canada, 

HARRINGTON,  Elliott  D.*  (M  1932;  A  1930), 
Engr*  (for  mail).  Air  Cond.  Dept.,  Commercial 
Kngrff,  DJv,,  General  Electric  Co.,  1  River  Rd., 
and  1680  Wendell  Ave»,  Schenectady,  N.  Y. 

HARRIS,  J«$a©  B.  (M  1018} ,  Pres.  (for  mail), 
Rose  &  Harris  Ena.»  Inc.,  416  Essex  Bldg.,  and 
3620  Colfax  Ave.  S.,  Minneapolis,  Minn. 

HART-BAKER,  H«my  W.  (M  1018),  Director 
(for  mail),  Merritt,  Ltd,,  8  French  Bund,  and 
87  Rte  Rene  Delastre,  Shanghai,  China. 

HART,  Harrv  M,*  (M  1012),  (PmidtntW 
Jkfomto),  (Fr<as,»  1916;  1st  Vke-Pres,,  1915; 
Council,  10144917),  Pres.  (for  mail),  L.  H, 
Prentice  Co.,  1048  Van  Buren  St.,  and  6409 

(    Wlntorop  Ave.,  Chicago,  111. 

HARTMAN,  Fred  Stewart  (A  1938),  Dist,  Mgr., 
Industrial  Dept-  (for  mail) ,  General  Electric  Co,, 
170  Lexington  Ave.»  New  York,  N,  Y,,  and  168 
Montckir  Ave.,  Montdalr,  N.  J. 


19 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


HARTMAN,  John  M.  (M  1927),  Engr.  (for  mail), 
Kewanee  Boiler  Corp.,  and  719  Henry  St., 
Kewanee,  III. 

HARTWEIN,  Charles  E.  (M  1933),  Supervisor, 
House  Htg.  Dept.,  St.  Louis  County  Gas  Co., 
231  W.  Lockwood,  Webster  Groves,  and  (for 
mail),  6271  Magnolia  Ave.,  St.  Louis,  Mo. 

HARTWELL,  Joseph  C.  (M  1922),  (for  mail), 
Hartwell  Co.,  Inc.,  87  Weybosset  St.,  and  10 
Freeman  Pkwy.,  Providence,  R.  I. 

HARVEY,  Alexander  D.  (A  1928;  J  1925),  Nash 
Engrg.  Co.,  South  Norwalk,  Conn. 

HARVEY,  Lyle  C.  (M  1928),  Vice-Pres.  (for  mail), 
Bryant  Heater  £  Mfg.  Co.,  17825  St.  Clair  Ave., 
and  3388  Glencarin  RdM  Cleveland,  Ohio. 

HASHAGEN,  John  B.  (M  1930),  121  Manhattan 
Ave.,  Jersey  City,  N.  J. 

HASLETT,  Henry  M.  (5  1935),  950  Lombard 
Ave.,  St.  Paul,  Minn. 

HATEAU,  William  M.  (J  1934),  Draftsman  and 
Student,  Sherron  Metallic  Corp.,  1201  Flushing 
Ave.,  Brooklyn,  and  (for  mail),  1530  Sheridan 
Ave.,  New  York,  N.  Y. 

HATTIS,  Robert  E.  (M  1926),  Consulting  Engr. 
(for  mail),  ISO  N.  Michigan  Ave.,  and  4251  N. 
Mozart  St.,  Chicago,  111. 

HAUAN,  Merlin  J.  (M  1933),  Consulting  Engr., 
3412-16th  S.,  Seattle,  Wash, 

HAUPT,  Howard  F.  (A  1929),  614  E.  Beaumont 
Ave,,  Milwaukee,  Wis. 

HAUSS,  Charles  F.*  (Charter  Member;  Life. 
Member),  Via  Gioberti  No.  2,  Milano,  Italy. 

HAYDEN,  Carl  F.  (A  1930),  Br.  Mgr.  (for  mail), 
Barber-Colman  Co.,  221  N.  LaSalle  St.,  Chicago, 
and  2227  Ewing  Ave.,  Evanston,  111. 

HAYES,  James  J.  (M  1920),  Sales  Engr.  <for 
mail),  Stannard  Power  Equipment  Co.,  53  W, 
Jackson  Blvd.,  Room  925,  and  7443  Jeffery  Ave., 
Chicago,  111. 

HAYES,  John  J.  (A  1933),  Auburn  Stoker  Sales 
Corp.,  40(J  N.  Wells  St.,  Chicago,  and  (for  mail), 
918  Michigan  Ave.,  Evanston,  111. 

HAYES,  Joseph  G.  (M  1908),  Pres.  and  Engr. 
(for  mail),  Hayea  Bros..  Inc.,  236  W.  Vermont 
St.,  and  2849  N.  Capitol  Ave.,  Indianapolis,  Ind, 

HAYMAN,  A.  Eugene,  Jr.,  (J  1935;  5  1930), 
2500  Washington  St.,  Wilmington,  Del. 

HAYNES,  Charles  V,  (M  1917),  (Presidential 
Member),  (Pres.,  1934;  1st  Vice-Prea.,  1933;  2nd 
Vice-Pres,,  1032;  Council,  102(5-1921),  11)32-1934), 
Vice-Pres.,  Hoffman  Specialty  Co.,  500  Fifth 
Ave.,  Room  3324,  New  York,  N.  Y.,  and  (for 
mail),  115  Llanfair  RcL,  Ardmore,  Mont.  Co.,  Pa. 

HAYTER,  Bruce  (M  1934),  Chief  Engr.,  Institute 
of  Thermal  Research  (for  mail),  American 
Radiator  Co.,  (575  Bronx  River  Rd.,  Yonkers, 
and  49  Carman  Rd.,  Scarsdalc,  N.  Y. 

HAYWARD,  Ralph  B.  (M  1009),  Prea.  (for  mail), 
R.  B.  Haywurcl  Co.,  1714  Sheffield  Ave,,  Chicago, 
and  201  S.  Stone  Ave.,  LaGrange,  III. 

HEARD,  John  A.  E.  (J  1930),  Carrier  Engrg.  Co., 
Ltd.,  Sardar  Sujan  Singh  Block,  Connaught  PL, 
New  Delhi,  India,  and  (for  maiU,  "Lyntan"  28 
Leighcliff  Rd,,  Leighon-Sea,  Eaaex,  England. 

HEARD,  Roderick  G.  (A  1933),  Asat.  to  Mgr,, 
Fuel  Oil  Dept.  (for  mail),  Imperial.  Oil,  Ltd., 
50  Church  St.,  and  12  Huntley  St.,  Toronto, 
Ont.,  Canada. 

HEATH,  William  R.  (M  1031),  Asat.  Chief  Engr., 
Buffalo  Forge  Co.,  400  Broadway,  and  (for  mall), 
119  Wingate  Ave,,  Buffalo,  N.  Y. 

HBBERUNG,  G.  W.  (A  1934),  Box  115,  Wayzata, 
Minn, 

HBOHT,  Frank  H.  (M  1930),  Saks  Engr.  (for 
mall),  B»  F*  Sturtevant  Co.,  2635  Koppers  Bidg., 
and  1467  Barnesdale  St.,  Pittsburgh,  Pa, 

HECKEL.  E.  F.  (M  1918),  Vice-Pres.  and  Gen, 

Mgr.,  Chicago  Dist.  (for  mall).  Carrier  Engrg. 

Corp..  180  N.  Michigan  Ave.,  Chicago,  and  314 

Cuttriss  PL,  Park  Ridge,  111. 

HEDGES,  H.  Berkley  (M  1019),  1021  Park  Lane, 

Plainfield,  N.  J, 

HB0LBY,  Park  S.  (M  1023),  Park  S.  Hedley  Co,, 
Curtiss  Bldg,,  Delaware  &t  Tapper,  Buffalo, 
N,  Y. 


HEEBNER,  Walter  M.  (M  1922^,  Sales  Engr., 
Warren  Webster  &  Co.,  470  Fourth  Ave.,  New 
York,  N.  Y.,  and  (for  mail),  282  Highwood  Ave., 
Teaneck,  N.  J. 

HEIBEL,  Walter  E,  (M  1917),  Dist.  Mgr.  (for 
mail),  Aerofin  Corp.,  11  West  42nd  St.,  New 
York,  N.  Y.,  and  Old  Greenwich,  Conn. 

HEILMAN,  Russell  H.*  (M  1023),  Senior 
Industrial  Fellow  (for  mail),  Mellon  Institute, 
and  5G37  Wilkins  Ave.,  Pittsburgh,  Pa. 

HELBURN,  I.  B.  (M  1929;  J  1927),  Junior  Assoc. 
(for  mail),  Wyman  Engrg.,  Chamber  of  Com- 
merce Bldg.,  and  700  Clialfonte  PL,  Apt.  17, 
Cincinnati,  Ohio. 

HELLSTROM,  John  (A  1929),  Vicc-Prcs.  (for 
mail),  American  Air  Filter  Co.,  215  Central 
Ave.,  Louisville,  and  Anchorage,  Ky, 

HENDRIGKSON,  Harold  M.  (M  1983),  Mech. 
and  Refrigeration  Engr.,  M.  J.  Hauan,  Con- 
sulting Engr.,  324-1411  Fourth  Ave.,  Bldg.,  and 
(for  mail),  7328  Earl  Ave.  N.W.,  Seattle,  Wash. 

HENDRICKSON,  John  J.  (.-1  1932),  Prod.  Engr, 
(for  mail),  Bryant  Heater  &  Mfg.  Co.,  17825 
St.  Clair  Ave.,  Cleveland,  and  1475  Gcnessee 
Rd.,  South  Euclid,  Ohio. 

HENION,  Hudson  D.  (A  1923),  Sales  Mgr.  (for 
mail),  C.  A.  Dunham  Co.,  Ltd.,  1523  Davenport 
Rd.,  and  45  Ridge  Dr.,  Toronto,  Ont.,  Canada. 

HENRY,  Alexander  S.t  Jr.  (M  1930),  300  Central 
Park  West,  New  York,  N.  Y. 

HERENDEEN,  Frederick  W.  (At  1020),  The 
Institute  of  Boiler  and  Radiator  Mfrs.,  29 
Seneca  St.,  Geneva,  N.  Y. 

HERKIMER,  Herbert  (M  1934),  Director  (for 
mail),  Herfcimcr  Inst.  of  Refrigeration,  181J) 
Broadway,  and  25  Central  Park  West,  New 
York,  N.  Y. 

HERLIHY,  Jeremiah  J.  (Life  Member;  M  1914), 
Pres.  (for  mail),  J.  J.  Herlihy,  Inc.,  810  W» 
Congress  St.,  and  3(534  N.  Keeler  Ave.,  Chicago, 
111. 

HERRIOK,  Daniel  A,  (M  1923),  Gen.  M«r,  (for 
mail),  Julian  d'Kete  Co.,  6  Spice  St.  (Charles- 
town  Dist.),  Boston,  and  27  Agassi/  St.,  Cam- 
bridge, Mass. 

HERRIOK,  Leo  (M  1935),  Mgr,  (for  mail),  Crane 
Co,,  and  323  Greenwood  Ave,,  Ft.  Smith,  Ark, 

HERRING,  Edgar  CM  1010),  Chairman  and 
Governing  Director  (for  mail),  J.  Jeffreys  &  Co., 
Ltd,,  Barrons  PL,  Waterloo  Rd.,  London  S.E., 
and  "Kenia"  Keswick  Rd,,  Putney,  London 
S.W.,  England. 

HERRMANN,  Harold  C.  (J  1935;  S  1032). 
Instructor,  Milwaukee  Vocational  School,  and 
(for  mail),  4523  North  22ncl  St.,  Milwaukee,  Wis, 

HERTY,  Frank  B.  (M  1933),  House  Htg,  vSup- 
erviaor,  Brooklyn  Union  Ga@  Co.,  180  Remaen 
St.,  and  (for  mail),  50  Bast  18th  St.,  Brooklyn, 
N.  Y. 

HERTZLE&,  John  R,*  (J  W28),  Air  Cond.  Saks 
Engr.  (for  mail),  York  Ice  Machinery  Corp., 
42nd  St.  and  Second  Ave,,  Brooklyn,  and  1 
University  Pl,»  New  York,  N,  V, 

HESS,  David  K.  (J  1935;  .$'  HUM,  Student  (for 
mail),  827  Mendota  Court.  Mttdtson,  Wia,,  and 
Hess  Warming  &  Vtg.  Co,,  mi-HJ27  S.  Western 
Ave,,  Chicago,  111. 

HESTER,  Thomas  J.  (M  1010),  Vice-Prew-Treas. 
(for  mall),  Heater- Bradley  Co.,  283$  Washington 
Blvd.,  and  67  Aberdeen  PL,  St.  Louis,  Mo, 

HEXAMER,  Harry  D,  (M  1931),  ftilei  Kngr.  (for 
mail),  Kxcelso  Products  Corp..  rtft  Clyde  Ave.» 
and  103  E.  Delavan  Ave.,  Buffalo,  N,  Y* 

HEYDON,  Charles  G.  (A  1923),  Mgr.  Sal««  of 
Western  Div.,  Wright-Austin  Co.,  315  W.  Wood- 
bridge  St.,  and  (for  mail),  2681  Nebraska, 
Detroit,  Mich, 

KftBBS,  Frank  G.  (M  1017),  Salesman,  The  H,  B. 
Smith  Co.,  2209  Chestnut  St.,  and  (for  mail), 
846  North  85th  St.,  Philadelphia,  Ft. 

HICKEY,  Daniel  W.  (A  1931),  278  W.  Fourth  St., 
St.  Paul,  Minn. 

HICK.EY.  J%m0a  W.  (J  1035;  3  1982),  P.  O.  Box 
245,  Carnegie  Institute  of  Technology*  Pitts* 
burgh*  Pa, 


20 


ROLL  OF  MEMBERSHIP 


HIERS,  Charles  R.  (M  1929;  A  1929;  J  1927), 
Mgr.,  Minneapolis-Honeywell  Regulator  Co., 
801  Second  Ave.,  New  York,  and  (for  mail), 
45-18-258th  St.,  Great  Neck,  L,  L,  N.  Y. 

HIGGINS,  Thomas  J.  (M  1927;  A  1927;  J  1923), 
P.  O.  Box  17,  East  Dedham,  Mass. 

HILDEBRANDT,  Henry  A.  (M  1918),  Supt.  of 
Bldgs.  and  Grounds,  University  of  Minnesota, 
and  (for  mail),  708  Sixth  Ave.  S.,  Minneapolis, 
Minn. 

HILL,  Dr.  E.  Vernon*  (M  1914;  A  1912),  (Presi- 
dential Member),  (Pres.,  1920;  1st  Vice-Pres., 
1919;  2nd  Vice-Pres.,  1918;  Council,  1915-1921), 
Pres.  (for  mail),  E.  Vernon  Hill  Co.,  121  N.  Clark 
St.,  and  1120  Farwell  Ave.,  Chicago,  111. 

HILL,  Fred  M.  (M  1930),  225  East  Ave.  39, 
Los  Angeles,  Calif. 

BILLIARD,  Charles  E.  (M  1932;  J  1927),  Htg. 
and  Vtg.  Engr.  (for  mail),  E.  C.  Hilliard  Co.,  27 
B  St.,  South  Boston,  and  1301  Washington  St., 
South  Braintree,  Mass. 

HILLS,  Arthur  H.  (M  1924),  Mgr.  (for  mail), 
Sarco  Canada,  Ltd.,  725-G  Federal  Bldg.,  83 
Richmond  St.  W.,  Toronto,  Ont.,  Canada. 

HINCKLEY,  Harlan  B.  (A  1934),  Engr.,  Custo- 
dian, Board  of  Education,  S510  S.  Green  St.,  and 
(for  mail),  0933  Princeton  Ave.,  Chicago,  111. 

HINKLE,  Edwin  C.  (Life  Member;  M  1911), 
170  N.  Franklin  St.,  Hempstead,  L.  L,  N.  Y. 

HINRICHSEN,  A.  F.  (M  1928),  Pres-Treas.  (for 
mail),  A,  F.  Hinrichsen,  Inc.,  50  Church  St., 
New  York,  N.  Y.,  and  Mountain  Lakes,  N.  J. 

HIRES,  J.  Edgar  (M  1927),  Pres.  (for  mail), 
Hires,  Castner  &  Harris,  Inc.,  200  South  24th 
St.,  Philadelphia,  and  107  Linwood  Ave., 
Ardmore,  Pa. 

HIRSCHMAN,  William  F.  (M  1929),  Pres.  and 
Chief  Engr.,  W.  F.  Hirschman  Co.,  Inc.,  220 
Delaware  Ave,,  Buffalo,  and  (for  mail),  105 
Le  Brun  Circle,  Eggertsville,  N,  Y, 

HITCHCOCK,  Paul  C.  (M  1931),  Partner  (for 
mail),  Burlingame  &  Hitchcock,  Consulting 
Engrs.,  520  Sexton  Blelg.,  and  4939  Girard  Ave. 
S.,  Minneapolis,  Minn. 

HJBRPE,  Clarence  A,,  Jr.  (J  1931),  73  Arch  St., 
New  Britain,  Conn. 

HOBBS,  J,  Clarenqe  (M  1920),  60  Wood  St., 
Painesvillc,  Ohio. 

HOCHULJU  Henry  W.  (M  1925),  Sales  Engr,, 
National  Radiator  Corp.,  55  West  42nd  St.,  New 
York,  N.  Y.,  and  (for  mail),  113  Chester  Ave., 
Bloomlield,  N.  J. 

HODEAUX,  W,  L.  (M  1931),  Owner  (for  mail), 
W,  L.  Hodeaux  Plbg.  &  Htg,  Co.,  216-17  N, 
fclagler  Dr.,  and  310-lOth  St.,  West  Palm  Beach, 
Fla. 

H0DGDQN,  Harry  A.  (Ad  1919),  153  Norfolk  St., 
Wollaaltm,  Mass. 

HODGE,  William  B,  (M  1034),  VIcc-Prea.,  Parka 
Cramer  Co.,  and  (for  mail),  F,  O.  Box  1234, 
Charlotte,  N.  C. 

HOFFMAN,  Charles  S,  (M  1924),  Vice-Pres.  (for 
mail),  Baker  Smith  &  Co.,  Inc.,  570  Greenwich 
St.  and  75  Central  Park  W.,  New  York,  N,  Y. 

HOFFMAN,  James  D,*  (M  1903),  (Presidential 
Umber),  (Pres.,  1D10:  1st  Vice-Pres.,  11)08; 
Board  of  Governors,  1911-11)12),  Prof,  of  Practi- 
cal Mechanics,  Head  of  Dept.,  Director  of 
Practical  Mech.  Lab.  (for  mall),  Purdue  Uni- 
versity, and  828  University  St.,  W.  Lafayette, 
Ind, 

HOFT,  Paul  J.  (M  1025;  A  1924),  (for  mail),  245 
8.  Eighth  St.,  and  1119  Wyoming  Ave,,  Phila- 
delphia, Piw 

HOOAN,  Edward  L,*  (M  1911),  Consulting  Engr, 
(for  mall),  American  Blower  Co,,  6000  Rusaell 
St.,  and  700  Sewatd  Ave,,  Detroit,  Mich, 
HOLBEOOK,  Frank  MJ*  (M  1023),  Enfir,,  62 

Walnut  Crescent  Montclair,  N.  J. 
HOIXA0AY,  William  L.  (A  1988),  Mgr.,  Engrg, 
Dept,  (for  mailj,  The  George  Belaey  Co,,  Ltd., 
1001  S.  Hope  St.,  LOB  Angeles,  and  110  Loma 
Alt*.  Dr,.  Altadena,  Calif, 
HQLUSTER,   B,    Wallace    (J    1931),    Owner, 
HoUUter'i*  21  Bay  St.,   Glens  Falls,  and  (for 
mall),  88  Oak  St.,  Hudson  Falls,  N.  Y, 


HOLLISTER,  Norman  A.  (M  1933),  Repr.,  The 
Trane  Co.,  122  East  42nd  St.,  New  York,  and 
(for  mail),  7101  Colonial  Rd.,  Brooklyn,  N.  Y.  ' 

HOLMES,  Richard  E.  (J  1934),  Engr.,  Air  Cond., 
Westinghouse  Elec.  &  Mfg.  Co.,  E.  Springfield, 
and  (for  mail),  11  Bushwick  St.,  Springfield, 
Mass, 

HOLT,  James  (M  1933),  Asst.  Prof,  (for  mail), 
Massachusetts  Institute  of  Technology,  Cam- 
bridge, and  1062  Massachusetts  Ave.,  Lexington, 
Mass. 

HOLTON,  John  H.  (M  1927),  Asst.  Chief  Engr. 
(for  mail),  Carrier  Engrg.  Corp.,  850  Freling- 
huysen  Ave.,  Newark,  and  5  Mountain  Ave., 
Maplewood,  N.  J. 

HOOD,  O.  P.  (Honorary  Member  1929),  (for  mail), 
Chief,  Technologic  Br.,  U.  S.  Bureau  of  Mines, 
17th  and  F  St.  N.W.,  and  1831  Irving  St.  N.W., 
Washington,  D.  C. 

HOOPER,  Vernon  F.  (M  1929),  Gl  Eastern  Ave., 
Ossining,  N.  Y. 

HOPPER,  Garnet  H.  (M  1923),  Engr.,  Taylor 
Forbes,  Ltd.,  1088  King  St.  W.,  and  (for  mail), 
19  Brummel  Ave,,  Toronto,  Ont.,  Canada. 

HOPSON,  William  T.  (Life  Member;  M  1915), 
The  Hopson  &  Chapin  Mfg.  Co.,  New  London, 
Conn. 

HORNUNG,  John,  C.  (M  1914),  Engr.  (for  mail), 
Central  Heat  Appliances,  343  S.  Dearborn  St., 
Chicago.,  and  854  Bluff  St.,  Glencoe,  HI. 

HORTON,  Homer  F.  (M  1925),  Sales  Repr.  (for 
mail),  National  Regulator  Co.,  2301  Knox  Ave., 
Chicago,  111. 

HOSHALL,  Robert  H.  (M  1930),  Associate  (for 
mail),  Thos.  H.  Allen,  Consulting  Engr.,  65 
McCall  St.,  and  789  N.  Evergreen  St.,  Memphis, 
Tenn. 

HQSKING,  Homer  L.  (M  1930),  Br,  Sales  Mgr, 
(for  mail),  Pacific  Steel  Boiler  Corp.,  370  Lexing- 
ton Ave.,  New  York,  and  5  Church  Lane, 
Scarsdale,  N,  Y. 

HOSTERMAN,  Charles  O.  (M  1924),  Supt.,  The 
McMurrer  Co.,  303  Congress  St.,  Boston,  and 
(for  mail),  25  Bateswell  Rd.,  Dorchester,  Mass. 

HOTCHKISS*  Charles  H.  B.  (M  1927),  Heating 
and  Ventilating,  US  Lafayette  St.,  New  York, 
N,  Y. 

HOUGHTEN,  Ferry  C.*  (M  1921),  Director  (for 
mail),  Research  Lab.,  A.S.H.V.E.  Bureau  of 
Mines,  4800  Forbes  St.,  and  1136  Murray  Hill 
Ave.,  Pittsburgh,  Pa. 

HOULISTON,  G.  Baillie  (A  1928),  Secy,  (for 
mail),  W.  C.  Green  Co.,  704  Race  St.,  Cincinnati, 
Ohio,  and  112  Forest  Ave.,  Ft.  Thomas,  Ky. 

HOWARD,  Ed&ar  (S  1936),  2721  Blaisdell  Ave., 
Minneapolis,  Minn. 

HOWATT,  John*  (M  1915),  (1st  Vice-Pres., 
1934:  2nd  Vice-Pres.,  1933;  Council,  1927-1934), 
Chiet  Engr,  (for  mail),  Board  of  Education,  228 
N.  LaSalle  St.,  and  4940  East  End  Ave.,  Chicago, 
III 

HGWELL,  Frank  B.  (M  1920),  Tech.  Advisor 
(for  mail),  American  Radiator  Co.,  40  West  40th 
St,  and  15  Central  Park  W.,  New  York,  N.  Y. 

HQWELL,  Lloyd  (M  1915),  Htg.,  Vtg.  and  Air 
Cond.  Engr,,  Peoples  Gaa  Bldg.,  122  S.  Michigan 
Ave.,  and  (for  mall),  7601  Yates  Ave.,  Chicago, 
III, 

HOWLETT*  Ira  G.  (S  1934),  2132  N,  PonshUl 
Ave,,  Oklahoma  City,  Okla, 

HOYT,  Charles  W,  (A  1931),  Pres-Treae.  (for 
mail),  Wolverine  Htg.  and  Vtg,  Equip,  Co.,  80 
Boylston  St,  Boston,  and  45  Thaxter  Rd,, 
Newtonville,  Maas* 

HOYT,  Leroy  W,  (M  1930),  N.  Stamford  Ave,, 
Stamford,  Conn. 

HUBBARD,  George  Wallace*  (M  1011),  Chief 
Mech,  Engr.  (for  mail),  Graham,  Anderson, 
Probst  &  White,  1417  Railway  Exchange, 
Chicago,  and  710  Bonnie  Brae,  River  Forest,  111. 

HUGH,  A,  J.  (M  1919),  Secy-Treas,  (for  mail) , 
Central  Supply  Co.,  312  S.  Third  St,  and  4087 
Harriet  Ave.,  Minneapolis,  Minn.  ' 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


HUCKER,  Joseph  H.  (M  1921),  Partner,  Hucker- 
Pryibil  Co,,  1700  Walnut  St.,  Philadelphia,  and 
(for  mail),  715  Stanbridge  St.,  Norristown,  Pa. 

HUDSON,  Robert  A.  (M  1934),  Consulting  Engr. 
(for  mail),  Hunter  &  Hudson,  41  Slitter  St., 
R.  710,  and  2850  Union  St.,  San  Francisco, 
Calif. 

HUETTNER,  Henry  F.  (S  1934),  124  Jerusalem 
Ave.,  Hicksville,  L.  L,  N.  Y.,  and  (for  mail), 
Box  388,  Carnegie  Tech.,  Pittsburgh,  Pa. 

HUFFAKER,  Herbert  B.  (ill  1933),  209  N.  Tenth 
St.,  Council  Bluffs,  Iowa. 

HUGHES,  Charles  E.  (S  1934),  140  Beacon  Ave., 
New  Haven,  Conn. 

HULL,  Harry  B.  (M  1931),  Mgr.,  Research 
Engr.,  Frigidaire  Corp.,  and  (for  mail),  P.  O. 
Box  671,  Dayton,  Ohio. 

HUMPHREY,  Dwight  E.*  (M  1921),  Htg.  and 
Vtg.  Engr.,  Goodyear  Tire  &  Rubber  Co.,  1144 
E.  Market  St.,  Akron,  and  (for  mail),  2499  Sixth 
St.,  Cuyahoga  Falls,  Ohio. 

HUMPHREYS,  Clark  M.  (M  1931),  Asst.  Prof,  of 
Mech.  Engr.  (for  mail),  Carnegie  Institute  of 
Technology,  Schenley  Park,  and  1934  Remington 
Dr.,  Pittsburgh,  Pa. 

HUNGER,  Robert  F.  (M  1927),  Associate  Dist. 
Mgr.  (for  mail),  Buffalo  Forge  Co.,  703  Cunard 
Bldg.,  and  4618  Chester  Ave.,  Philadelphia,  Pa, 

HUNGERFORD,  Leo  (M  1930),  Pres.  (for  mail), 
Pacific  Elec.  &  Mech.  Co.,  Inc.,  524  Loew's  State 
Bldg.,  and  105  N.  Berendo  St.,  Los  Angeles, 
Calif. 

HUNT,  Noel  P.  (M  1934),  Managing  Dir.  (for 
mail),  Carrier  Australasia,  Ltd.,  50  Hunter  St., 
and  52  Lang  Rd.,  Centennial  Park.,  Sydney, 
N.S.W..  Australia. 

HUNZIKER,  Chester  E.  (A  1934),  Br.  Mgr.  (for 
mail),  American  Blower  Corp,,  331  State  St., 
Schenectady,  and  422  Reynolds  St.,  Scotia,  N.  Y. 

HUSKY,  S.  T.  (S  1934),  710  Monnett  St.,  Nor- 
man, Okla. 

HUST,  Carl  E.  (At  1932),  Htg.  Engr.  (for  mail), 
The  Union  Gas  &  Elec.  Co.,  Room  1008,  Fourth 
and  Main  Sts.,  and  Hillcrest  Apts.,  15  Mason  St., 
Cincinnati,  Ohio. 

HUSTOEL,  Arnold  M.  (A  1030),  12023  N.  Ballou 
St.,  Chicago,  III. 

HUTGHINS,  William  II,  (M  1934),  Chief  Engr,, 
Delco  Appliance  Div.,  General  Motors  Corp,, 
and  (for  mail),  88  Magee  St.,  Rochester,  N.  Y. 

HUTZEL,  Hugo  F.  (M  1918),  Mgr.,  Erskine 
Radiator  Div.  (for  mail),  Chase  Brass  &  Copper 
Co,,  Inc.,  and  72  Hewlett  St.,  Waterbury,  Conn. 

HVOSLEF,  Fredrik  W.  {M  19U1;  A  1921),  Htg. 
Research  Kngr,  (for  mail),  Kohler  Co,,  and  52tt 
Autlubon  Rd.,  Kohler,  Wia. 

HYDE,  Lawrence  L,  (J  1934),  Gen.  Mgr,,  M.  J. 
O'Ncil,  and  (for  mail),  M  S.  Cretin  St.,  St.  Paul* 
Minn. 

HYMAN,  Wallace  M.  (M  1920),  Pres.  (for  mail), 
Rots  &  O'Donovan*  Inc.,  12  West  21st  St.,  and 
23  West  73rcl  St.,  New  York,  N.  V, 

HYNES,  Lee  F,*  (M  1010),  (for  mail),  240  Cherry 
St.,  Philadelphia,  ?a,,  and  Haddoafield,  N,  J. 

I 

ICKERINGILL,  J.  a  (AT  1923),  Engr..  Spencer 

Heater  Co.,  2504,  N.  Broad  St.,  and  (for  mail), 

23S  Rector  St.,  Philadelphia,  Pa. 
ING  ALLS,  Frederick  D.  B.  (M  1906),  Consulting 

Engr,,  1  Hopkins  St.,  Reading,  Mass. 
INGELS,  Margaret*  (M  1928;  J  1018),  Mech. 

Engr,    (for  mail),   Carrier   Engrg.    Corp.,   8&0 

Freunghuysen  Ave.,  Newark*,  and  Hotel  East 

Orange,  Bast  Orange,  N.  J. 
INGOLD,  John  W.  (7  1035;  S  1938),  3038  Perry- 

ville  Ave.  N.S.,  Pittsburgh,  Pa. 
INNIS,  Helm  R*  (M  1921;  J  1918),  Urgent, 

W.Va. 
I8SERTELL,   H«nry  G,*   (M  m3;   A    1912), 

Consulting  Engr.,  81  Park  Terrace  W.,  N<sw 

York,  N.  Y, 


JACKSON,  Alton  B.  (M  1932),  15  Hcrrick  St., 

Winchester,  Mass. 
JACKSON,  Charles  H.  (M  1923),  Vice-Pres.  (for 

mail),  Blower  Application  Co.,  918  N.  Fourth  St., 

and  2700  N.  Farwell  Ave.,  Milwaukee,  Wis. 
JACKSON,   Marshall   S.    (M   19110,    Rcpr.    (for 

mail),  Powers  Regulator  Co.,  250  Delaware  Ave., 

and  10S  Larchniont  Rd.,  Buffalo,  N.  Y. 
JACOBUS,  Dr.  David  S    (Lifc  Member;  Al  1910), 

The    Babcock  &   Wilcox    Co.,   85    Liberty   St., 

New  York,  N.  Y. 
JALONACK,  Irvviii  G.  (A  1933;  6'  1930),  Knsrg. 

Mgr.  (for  mail),  c/o  A,  L.  Hart,  82  Railroad  Ave., 

and  15  South  St.,  Patchotfue,  N.  Y. 
JAMES,  Hamilton  R.  (A/  1931),  Service  Kquip. 

Engr.,  United  Engineers  &  Constructors,   Inc., 

1401  Arch  St.,  Philadelphia,  and  (for  mail),  55 

W.  Drexcl  Ave.,  Lansdowne,  Pa. 
JAMES,  John  W  *  (7  1933),  2223  Grand  Blvd., 

Schenectady,  N.  Y. 
JANET,   Harry  L.   (Al   19130),   KHRI.    (for  mail), 

Carrier  Engrg.  Corp.,  Chrysler  BUltf.,  New  York, 

and  688  DecaturSt.,  Brooklyn,  N.  Y. 
JARDINE,  Douftlas  C.  (M  102');  A   102t>),  Pres. 

(for  mail),  Jardine  &  Knight  Plbg.  &  Ht«.  Co., 

312  N.  Custer  Ave.,  and  1512  K,  Platte  Ave., 

Colorado  Springs,  Colo. 
JARRATT,  Paul  R.  (.4  1031),  117  Fifth  Ave.  N., 

Nashville,  Temi, 
JELLETT,  Stewart  A.*  (Uonvmry  .Uembfr  1920), 

(Charter  Member;  Presidential  Alcmber),  (Pres., 

1805;  Board  of  MRW.»  1800-1800;  Secy.,  1808),' 

42(>  W.  Kllet  St.,  Germantown,  Philadelphia,  Pa. 
JENNEY,  Hu&h  B.  (.1  1033),  Sales  M«r.,  Domi- 
nion Radiator  &  Boiler  Co.,  Ltd.,  Royce  and 

Lansdowne  Avcs.,  Toronto,  Ont.,  Canada, 
JENNINGS,  Irving  0.  (At  1024),  Prow,  (for  mail). 

Nash  Engrg.  Co.,  and  138  Flax  Hill  Rd,,  South 

Norwulk,  Conn. 
JENNINGS,  Warren  G.  (A  1030),  Minueapolia- 

Honeyvvell    Regulator    Co,,    43    E.     Ohio   St., 

Chicago,  111. 
JENN1NS,  Henry  H.   (Life  Member;   M   1001), 

lf>    Grange    View,    Chapeltown     Rd.»     Leeds, 

JENJ5?ON(f  Joan  S.  (M  1012),  Consulting  Engr.  (for 

mail),  Neiler-kkh  &  Co.,  431  S.  Dearborn  St., 

and  1634  West  100th  St.,  Chicago,  111. 
JKPERT1NGER,   Richard  C.    (A    1034),    Gen, 

Mgr.   and   Engr.   (for  mail),   hyncromtitic  Air 

Cond,  Corp.,  1317  N.  Third  St.,  and  1028  W. 

Vienna  St.,  Milwaukee,  Wia. 
JOHN,  Victor  P.  (M  1031),  13(5  Herryman  Dr., 

Snydcr,  N,  Y. 
JOHNS,  Harold  B.*  (M  1028}  J  1027),  (for  mail), 

Peoples  Gas  Co.,  122  S.  Michigan  Av<a.,  Chicago, 

and  543  N.  Klmwood  Ave,,  Oak  Park,  111. 
JOHNSON,  Carl  W.  (Af  11)12).  Pres.  (for  mall), 

C.  W.  Johnson  Co.,  Inc.,  211  N,  Desplainoi  St., 

and  1800  Morse  Ave.»  Chicago.  III. 
JOHNSON*   Clarence  W.    (M   19^3;   A    1938; 

J  mi),  Br,  Mgr,  (for  mail),  Canadian  Sirocco 

Co.,    Ltd.,    030   Dorchester   Si,    W,,   and   3^3 

Dresden  Ave.,   Mt.    Royal,    Montreal,    P,   Q,, 

Canada, 
JOHNSON,  Edward  B,  (M  1019).  Sates  Engr,, 

Staten  lalaml  Supply  Co,,  and  (for  mail),  lf>4 

Wardwell  Ave,,  W,  New  Brighton,  N.  Y. 
JOHNSON,  Edftar  Jtt,  (M  1020),  Sales  Engr.  (for 

mail),  Buffalo  Forge  Co.,  490  Broadway,  and  103 

University  Ave.,  Buffalo,  N.  Y. 
JOHNSON,  H«te*  S.  (A  1083;  J  1027),  Buffalo 

Forge  Co,,  414  Standard  Bldg.,  Albany,  N.  Y, 
JOHNSON,  Leslie  0.  (J"  1930),  c/o  Y.M.CA,, 

White  Plains,  N,  Y. 
JOHNSON,  Louis  H.  (M  1031),  918  USftlte  Av«,, 

Minneapolis,  Minn, 
JOHNSTON,  HU&O  0»   (A   1984),   Coatraotor» 

Jleating  &  Ssmitary  Service,  Box  W2»  Wttllington 

0»L,  Canada. 
JOHNSTON,  J*  AmWter  (M  101»),  Ftrtner  tf<sr 

mull),  Carneal  Johnston  &  Wrtilit,  80S 

Bidg*.  Richmona,  Va* 


ROLL  OF  MEMBERSHIP 


JOHNSTON,  Robert  Elliott  (M.  1929;  A  1926), 
3342-33rd  Ave.  W.,  Vancouver,  B.  C.,  Canada. 

JOHNSTON,  William  H.  (M  1924),  306  East 
26th  St.,  New  York,  N.  Y. 

JONES,  Alfred  (M  1928),  Chief  Consulting  Engr. 
(for  mail),  Armstrong  Cork  Co.,  P.  O.  Box  540, 
402  President  Ave.,  Lancaster,  Pa. 

JONES,  Alfred  L.  (M  1926),  Plbg.  and  Htg. 
Contractors  (for  mail),  21  Church  St.,  Green- 
wich, and  Box  121,  Riverside,  Conn. 

JONES,  Bernard  G.  '(M  1928),  Mgr.  (for  mail), 
Acme  Fan  &  Blower  Co.,  Ltd.,  868  Arlington  St., 
and  542  Raglan  Rd.,  Winnipeg,  Man.,  Canada. 

JONES,  Charles  R.  (A  1928),  Jones  Supply  Co., 
Siloam  Springs,  Ark. 

JONES,  Edwin  (M  1933;  J  1924),  Box  582,  Tulsa, 
Okla. 

JONES,  Edwin  A.  (M  1919),  Chief  Engr.  (for 
mail),  L.  J.  Mueller  Furnace  Co.,  2001  W. 
Oklahoma,  and  4065  N.  Prospect,  Milwaukee, 
Wis. 

JONES,  Edwin  F.  (M  1923),  Consulting  Engr., 
420  New  York  Bldg.,  and  (for  mail),  220  Mont- 
rose  PL,  St.  Paul,  Minn. 

JONES,  Harold  L.  (M  1920),  Asst.  Supt.  (for 
mail),  W.  W.  Farrier  Co.,  44  Montgomery  St., 
Jersey  City,  and  11  Cambridge  Rd.,  Glen  Ridge, 

JONES,  Noel  W.  (/  1935;  5  1933),  1315  East  28th 
St.,  Minneapolis,  Minn. 

JONES,  Raymond  E.  (M  1919),  Asst.  Supervisor, 
Fuel  Oil  Sales,  Gulf  Refining  Co.,  1515  Locust 
St.,  Philadelphia,  Pa.,  and  (for  mail),  39  West 
End  Ave.,  Haddonfield,  N.  J. 

JONES,  William  T.  (M  1915),  (Presidential 
Member'),  (Pres,,  1933;  1st  Vice-Pres.,  1932;  2nd 
Vice-Pres.,  1931;  Council,  1925-1934),  Treas., 
Barnes  &  Jones,  128  Brookside  Ave.,  Jamaica 
Plain,  and  (for  mail),  1886  Beacon  St.,  Waban, 
Mass. 

JORDAN,  Lebcrt  E.  (A  1934),  Engr.  (for  mail), 
Minneapolis  Air  Conditioner  Co.,  1609  Hennepin 
Ave.,  and  122  Arthur  Ave,,  S.K.,  Minneapolis, 
Minn. 

JORDAN.  Richard  C.  (/  1935;  S  1933),  Air  Cond. 
Engr.,  Central  Supply  Co,,  312  S.  Third  St.,  and 
(for  mail),  2518  Grand  Ave.  S.,  Minneapolis, 

JOYCE,  Harry  B.  (M  1922),  Consulting  Engr.  (for 
mail),  810  Commerce  Bldg.,  and  501  Liberty  St, 
Erie,  Pa. 

JUNG,  John  S.  (M  1930;  A  1923),  Heating 
Contractor  (for  mail),  2409  W.  Greenfield  Ave,, 
and  1510  S.  Layton  Blvd,,  Milwaukee,  Wis. 

JtJTTNER,  Otto  J.  (M  1915),  Pres,  (for  mail), 
Tuttner  Htg.  Co.,  814  N.  Milwaukee  St.,  and  910 
E.  Wells  St.,  Milwaukee,  Wis. 


KAGEY,  I.  B.,  Jr-*  (/  1929),  Air  Cond,  Engr., 
Metropolitan  Life  Insurance  Co,,  1  Madison 
Ave.r  New  York,  N.  Y. 

XCAHAN,  Charles  (J  1935;  S  1038),  6  Marie  Ave,, 
Cambridge,  Mass, 

JCAHNSKY,  Alex  G,  (S  1934),  137QG  Durkec 
Ave,,  Cleveland,  Ohio, 

KAMMAN,  Arnold  R*  (A  192A;  J  1921),  (for 
mall),  Park  S.  Hedky  Co.,  304  Curtiss  Btdg,, 
Buffalo,  and  Wanakah,  Eric  Co.,  N,  Y, 

KAMPISH,  Nick  B,   (S  1984),  214  E,  Lincoln 

*  Ave.,  Roselle  Park,  N,  J,         -     ^    ^    - 

KAJPPEL,  George  W,  A.  (M  1921),  Pres.  and 
Treas.  (for  mall),  Camden  Heating  Co.,  Wilson 
Blvd.  and  Waldorf  Ave,,  Camden,  and  347  W. 
Kings  Highway,  Haddonfield,  N,  J. 

St  Louis  (J  1035;  S  1983),  00$  E.  Pearl 


*  Of  1918),  Chief  Engr.  (for 
mail),  Parks-Cramer  Co.,  970  Main  St,  Fitch- 
burg,  and  180  Prospect  St.,  North  teominster, 
Mass* 

KARTORIB,  V.  T.  (/  19S5j  3  1988),  Graduate 
A«it,  fa  Meeh,  Engr,,  Case  School  of  Applied 
Seteacft,  and  (for  mall),  2982  $ut  102nd  St., 
Cleveland,  Ohio. 


KASTNER,  George  G.   (J  1935;  5  1033),  Sales 

Engr.,  Schwerm  Air  Cond.   Corp.,  2303  Grand 

Concourse,  and  (for  mail),  654  East  226th  St., 

New  York,  N.  Y. 

KAUFMAN,  William  M.  (J  1935;  S  1933),  2875 

Sedgwick  Ave.,  New  York,  N.  Y. 
KEEFE,    Edmund   T.    (M    1931),   75   Pitts   St,, 

Boston,  Mass. 

KEENEY,    Frank   P.    (A    1915),   Pres.,    Keeney 
Publishing  Co.,  6  N.  Michigan  Ave.,  Chicago,  111. 
KEHM,   Horace  Stevens   (M  1928),   Pres.    (for 
mail),  Kehm  Bros.  Co.,  51  E.  Grand  Ave.,  and 
5510  Sheridan  Rd.,  Chicago,  111. 
KELBLE,  Frank  R.   (M  1928),  Vice-Pres.  and 
Mgr.   (for  mail),  Huff  man- Wolfe  Co.  of  Phila- 
delphia, 11  W.  Rittenhouse  St.,  Philadelphia,  and 
115  Roslyn  Ave.,  Glenside,  Pa. 
KELL,  Waldo  R.  (A  193i),  Sales  Engr.  (for  mail), 
The  Marley  Co.,  1915  Walnut  St.,  and  113  East 
69th  Terrace,  Kansas  City,  Mo. 
KELLEY,  James  J.  (A  1024),  Vice-Pres.  and  Gen. 
Mgr.   (for  mail),  Arthur  H.   Ballard,  Inc.,  535 
Commonwealth  Axre.,   Boston,  and  142  Gover- 
nors Ave.,  Medford,  Mass. 

KELLNER,  Day  C.  (S  1933),  (for  mail),  Carnegie 
Institute  of  Technology,  Pittsburgh,  Pa.,  and 
Cuba  City,  Wis. 

KELLOGG,  Alfred  (M  1916),  (Council,  1920- 
1921,  1923-1924),  Consulting  Engr.  (for  mail), 
585  Boylston  St.,  Boston,  and  C  Hawthorne  St., 
Belmont,  Mass. 

KELLY,  Charles  J.  (M  1931),  New  York  Repr., 
Jas.  P.  Marsh  Corp.,  551  Fifth  Ave.,  New  York, 
and  (for  mail),  162  Fairview  Ave.,  Jersey  City, 
N.  J. 

KELLY,   Hugh   (M  1927),  Mgr.   (for  mail),   H, 
Kelly   &    Co.,    Ltd.,    10041-101    A   Ave.,    and 
10235-124th  St.,  Edmonton,  Alberta,  Canada. 
KELLY,   John   G.    (A    1919),   374   Park  Ave., 

Yonkers,  N.  Y. 
KENDALL,   Edwin  H.    (A    1932;  J   1930),   309 

West  12th  St.,  Los  Angeles,  Calif. 

KENNEDY,  Maron   (J  1930),  Sales  Engr.   (for 

mail),  York  Ice  Machinery  Corp.,  5051  Santa  Fe 

Ave.,  and  2037  Bronson  Ave.,  Los  Angeles,  Calif. 

KENNEDY,  Owen  A.  (3  1933),  Carnegie  Institute 

of  Technology,  Pittsburgh,  Pa. 
KENNEDY,   Paul  V.    (S  1934),   105  Avon  St., 
New  Haven,  Conn.,  and  (for  mail),  4915  Forbes 
St.,  Pittsburgh,  Pa. 

KENT,  J.  King  (J  1928),  Pres.  (for  mail),  J.  King 
Kent  &  Co.,  Inc.,  0327  Clayton  Ave.,  and  5041 
Waterman,  St.  Louis,  Mo. 
KENT,  Laurence  F,  (A  1927;  J  1924),  Pres.  (for 
mail),  Moncrief  Furnace  Co.,  P.  O.  Box  1673, 
Atlanta,  and  R.  F.  D.  No.  2,  Smyrna,  Ga. 
KENWARD,  Stanley  B.  (J  1935;  5  1933),  45 

Fifth  Ave,,  Bay  Shore,  N.  Y. 
KEPLER,  Donald  A*  (S  1934),  Gibba  &  Cox,  Inc., 
11  Broadway,  New  York,  N.  Y.,  and  (for  mail), 
30  Maplewood  Ave.,  Maplewood,  N.  J. 
KEPLINGER,   William   L.    (M  1929),  Special 
Repr.    (for  mail),   Carrier   Kngrg.   Corp.,   408 
Chrysler  Bldg.,  New  York,  and  103  Sunset  Dr,f 
Hempatead,  L.  I.,  N.  Y. 

KEPPNER,    Harry   W*    (U   1930),    (for   mail), 
H,  W.  Keppner,  1310  South  56th  Ave.,  and  1245 
S.  Austin  Blvd.,  Cicero,  111. 
KERN,  Raymond  T-    (M   1927),   Chief  Engr,, 
Jenmaon  Co.,  Pitchburg,   and   (for  mail),   51 
Claflin  St.,  Leorainster,  Mass. 
KERSHAW,    Melville    O.    (M   1932;   A    1926; 
J  1921),  Vtg,  and  Air  Cond,  Engr.  (for  mail), 
E,  I,  Du  Pont  de  Nemours  &  Co.,  Wilmington* 
Del,  and  7313  North  21st  St.,  Philadelphia,  Pa. 
KEYES,  Robert  E.  (M  1918),  Chief  Engr.,  The 
Cooling  &  Air  Conditioning  Corp.,  24  Damon  St., 
Hyde  Park,  Boston,  Mass, 
KEYS,  Lee  Farls  (S  1934),  Asat.  Air  Cond.  Engr., 
Metropolitan  Life  Insurance  Co.,   1   Madison 
Ave.,  New  York,  N.  Y.,  and  (for  mail),  Route  1, 
Booc  866,  Phoenix,  Ariz. 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


KIGZALES,  Maurice  D.  (M  1935),  Assoc.  Mech. 

Engr.,    U.    S.    Army    Motion    Picture    Service, 

State,  War  and   Navy   Bldg.,   and    (for  mail). 

3000  Connecticut  Ave.  N.W.,  Washington,  D.  C. 
KIEFER,  Carl  J.  (M  1922),  Consulting  Engr.  (for 

mail),  918  Schmidt  Bldg.,  and  984  Lennox  PL, 

Cincinnati,  Ohio. 
KIEFER,  E.  J.,  Jr.  (A  1932;  J  1928),  Treas.  and 

Gen.  Mgr.  (for  mail),  H.  C.  Archibald  Co.,  8  S. 

Sixth  St.,  and  10S  N.  Sixth  St.,  Stroudsburg,  Pa. 
KIESLING,  Justin  A.  (M  1930),  Pres.  (for  mail), 

Robischimg-Kiesling,  Inc.,  4S4S  Main  St.,  P.  O. 

Box   1295,   and    1806   Holman   Ave.,    Houston, 

Texas. 
KILNER,  John  S.  (M  1929),  Owner  (for  mail), 

Kilner  Co.,  427  Stormfeltz-Loveley  Bldg.,  and 

1091  Seminole  Ave.,  Detroit,  Mich. 
KIMBALL,  Charles  W.  (M  1915),  Richard  D. 

Kimball  Co.,  6  Beacon  St.,  Boston,  Mass. 
KIMBALL,  Dwight  D.*  (M  190S),  (Presidential 

Member),    (Pres.,    1915;   2nd    Vice-Pres.,    1014; 

Board    of    Governors,    1912-1910),    Consulting 

Engr.  (for  mail),  205  East  42nd  St.,  and  307  East 

44th  St.,  New  York,  N.  Y. 
KING,  Roy  L.  (J  1935;,  5  1933),  2538  Clinton 

Ave.  S.,  Minneapolis,  Minn. 
KIPE,  J.  Morgan  (M  1919),  801  Homestead  Ave., 

Beechwood,  Del.  Co.,  Pa. 
KIRK,  Charles  D.   (M  1909),  Mgr.  (for  mail)* 

Chas.  D,  Kirk  Co.,  Sargent  &  Colleen,  and  774 

McMillan  Ave.,  Winnipeg,  Man.,  Canada. 
KIRKPATRICK,     Arthur    H.     (7    1931),     Ilg 

Electric  Vtg.  Co.,  415  Brainard  St.,  and   (for 

mail),  Webster  Hall  Hotel,  Detroit,  Mich. 
KIT  AURA,    Shigeyuki    (M   1918),    191    Shimo- 

ohsaki,  near  Tokyo,  Japan. 
KITCHELL,  Herbert  N.  (A   1926),  4528  Circle 

Ave.,  Cincinnati,  Ohio. 
KITCHEN,  Francis  A.  (A  1927;  J  1923),  Pres. 

(for  mail),  American  Wanning  &  Ventilating  Co., 

1514   Prospect  Ave,,   and    1711    Ivenyon   Rd., 

Cleveland,  Ohio. 
KITCHEN,  John  H,  (M  1906),  Owner  (for  mail), 

John  H,  Kitchen  Co.,  1016  Baltimore  Ave.,  and 

5015  Westwood  Terrace,  Kansas  City,  Mo. 
KLEIN,  Albert  (M  1920),  Managing  Director  (for 

mail),  Carrier  Lufttechnische  Gcscllschaft  m  b  H, 

Stuttgart,  Archivstrasse  14/16,  and  Stuttgart, 

Panoramastrasse  23,  Germany. 
KLEIN,  Edward  W.  (M  1917),  S.  E.  Diet.  Mgr, 

(for  mail),  Warren  Webster  &  Co.,  152  Nassau 

St.    N.W.,    and    456    Peachtree    Battle    Ave., 

Atlanta,  Ga. 
KLIE,  Walter  (M  1915),  Prea.  (for  mail),  The 

Smith  &  Oby  Co,,  0107  Carnegie  Ave.,  Cleveland, 

and  18411  S,  Woodland  Ave.,  Shaker  Heights, 

Ohio, 
KNIBB,  Alfred  E.  (M  1930),  Htg.  Engr,  (for  mail), 

1003  Maryland  Ave.,  and  9333  E.  Jefferson  Ave,, 

Detroit,  Mich, 
KNOPF,  Charles  (J  1935;  S  1933),  1201  Liberty 

Ave.,  Brooklyn,  N,  Y. 
KNOX,  James  R.  (U  1930),  Htg,  Engr.,  c/o 

Carmichael,  93  Arbroath  Rd,,  Dundee,  Scotland, 
KNUDTSON,  Carl  M.  (S  1935),  2410  Cromwell 

Dr.,  Minneapolis,  Minn, 
KOCH,   Harry  O.   (M  1916),  212  Centre  St., 

Tamaq.ua,  Pa. 
KOHLER,  Walter  J-T  Jr.  (A  1983),  Htg.  Sales 

Supervisor  (for  mall),  KohW  Co.,  and  005  W. 

Park  Lane,  Kohler,  Wia. 
KOLINSKY,  Max  D,  (S  1035),  418  Concord  St., 

St.  Paul,  Minn. 
KONZO,    Selchi*    (/    1032),    Special    Research 

Associate,  Engrg.  Experiment  Station.  University 

of  Illinois,  214  Mech.  Engr.  Lab.,  and  (For  mall), 

1108  W,  Stoughton  St.,  Urbana,  111. 
KO01STRA,  John  F.  (M  1033),  Engr.  (for  mail). 

Carrier  Bngrg.  Corp,,  748  E.  Washington,  and 

6063  Roy  St.,  Los  Angeles,  Calif. 
KORN,  Charles  B.  (M  1922),  Member  of  Firm, 

Reber«Korn  Co.,  817  Cumberland  St.,  and  (for 
.mail),  1022  S.  Eighth  St.,  Allentown,  Pa. 


KOTTCAMP,  Horace  A.   (M  1915),  Gen.  Mgr. 

(for  mail),  Kottcamp  Construction  Co.,  147  N. 

Second  St.,  and  527  Philadelphia  Ave.,  Chambers- 
burg,  Pa. 
KOZU,  Tamiichro  (U  1930),  No.  1701  Yonchome, 

Shimoochai,  Yodobashiku,  Tokyo,  Japan. 
KRAMIG,  Robert  E.,  Jr.   (A  1933),  Vice-Pres. 

and  Treas.  (for  mail),  R.  E.  Kramig  &  Co.,  222 

East    14th    St.,    Cincinnati,    and    51    Central 

Terrace,  Wyoming,  Ohio. 
KRATZ,  Alonzo  P.*   (M  1925),   Research  Prof. 

(for  mail),  Dept.  of  Mech.  Engrg.,  University  of 

Illinois,  and  1003  Douglas  Ave.,  Urbana,  111. 
KREISSL,  Hans  George  (M  1925),  Mgr.  "Vento" 

Dept.   (for  mail),  American  Radiator  Co.,  816 

S.   Michigan  Ave.,   Chicago,  and  408   Lee  St., 

Evanston,  111. 
KRIEBEL,  Arthur  E,  (M  1920),  Sales  En&r.  (for 

mail),  Haynes  Selling  Co.,  1518  Fairmount  Ave., 

Philadelphia,  and  Berwyn,  Chester  Co.,  Pa. 
KRUEGER,  James  I.  (M  1921),  Mfrs.  Repr.  (for 

mail),  357  Ninth  St.,  and  1920  Sacramento  St., 

San  Francisco,  Calif. 
KRUSE,  Robert  W.   (A   1930),  Krusc  Co.,  353 

West  IGth  PL,  Indianapolis,  Ind. 
KUEHN,  Walter  C.  (A  1933),  Kuelm  Htg.  &  Vtg. 

Co.,  915  Seventh  Ave.  S.,  Minneapolis,  Minn. 
KUEMPEL,  Leotx  L.  (/  1929),  3836  Vincent  Ave. 

S.,  Minneapolis,  Minn. 
KUHLMANN,  Rudolf  (A/  1928),  122  East  42nd 

St.,  New  York,  N.  Y. 
KUNS,  Joseph  W.   (S  1935),  2110  Hawthorne 

Ave.,  Minneapolis,  Minn. 
K.WAN,  I.  K.  (/U  1933),  Gen.  Mgr.,  China  Engrg. 

Co.,  30  Brenan  Rd.,  Shanghai,  China. 
KYLBERG,  V.  C.  (A  1934),  tK)7  Ridgewoud  Rd., 

Maplewood,  N.  J. 


LAGODZINSKI,  H.  J.  (A  1927;  J  1920),  8028 
N.  Tripp  Ave,,  Chicago,  III. 

LAN0AUER,  Leo  L,  (J  1932),  Mech.  Kn«r.  (for 
mail),  Kribs  &  Lamlauer,  Consulting  Kngrs., 
807  S.  W.  Life  iiklg.,  and  5707  Voluaco  St., 
Dallas.  Texas. 

LANDERS,  John  J.  (M  11)30;  A  1924:  J  1024), 
Sales  Engr,  (for  mail),  Pacific  Steel  Boiler  Corp., 
303  Crosby  Bldg.,  Buffalo,  and  Seneca  St., 
Kbcnezcr,  N.  Y. 

LANE,  0.  Dufly  (M  1934),  Supervisor,  The 
Servicing  Corp.  of  New  York,  101-10  Jamaica 
Ave.,  Jamaica,  and  (for  mail),  30*33-3Gth  St., 
Jackson  Heights,  N,  Y. 

LANGE,  Fred  F.  (A  1934),  Engr.  (for  mail), 
Mechanical  Service  Co.»  04 1  Northwestern  Bank 
Bldg.,  and  626a-28th  Ave.  S3.,  Minneapolis,  Minn. 

LANGBNBERG,  E,  B,  (M  1014),  (Council,  1926- 
1981),  Prea.  (for  mail),  Langenberg  Heating  Co., 
3800  W.  Pine  Blvd.,  and  6031  Snrijiht,  St.  Louis, 
Mo. 

LANNING,  E,  K.  (A  1927),  Warren  Webster  & 
Co,,  Camden,  N.  J. 

LANOU,  J,  Ernest  (M  1931),  Mgr.  (for  mail), 
F.  S,  Lanou  &  Son,  90  St.  Paul  St.,  and  48  Brookea 
Ave.,  Burlington,  Vt. 

LARSON,  GiMtu*  L.*  (M  1028),  (2nd  Vice-Pros,, 
1934;  Council,  1929-1034),  Prof,,  Steam  and  Gas 
Engrg.,  and  Chairman  of  Dept*  of  Mech,  Engrg., 
University  of  Wisconsin.  MadJaon,  Wis. 

LftSALVlA»  James  J,  (M  1930),  Air  Cond.  Engr., 
Frigidaire  Corp.  and  (for  mail),  Commodore 
Apts.,  522  Grand  Ave,,  Dayton,  Ohio, 

LAUTENSCHLAGER,  Ft«d  (M  1915),  Viefc- 
Pre8*Treaa.  (Offices  Kroeschell  Boiler  Co,,)  8253 
Kedsie  Ave.,  Chicago  (factories),  100  Eelchert 
Ct.,  Racine,  Wis»,  and  (for  mail)»  3840  Atta 
Vista  Terrace,  Chicago,  III 

LAW3LER.  Mutthew  M,  <J  1980),  Eeeident  Mgr. 
(for  mail),  Kitchen  Engrg,  Co»,  Inc**  1U  w. 


Washington  St.,  and  6900  N,  Aahtamd  Blvd., 
Chicago,  III. 

LAWTON,  Frank  a  (M  102S),  145  Euena  ViiU 
Ave.,  Hawthorne,  N.  J. 


ROLL  OF  MEMBERSHIP 


LEDNUM,  J.  Maynard  (M  1934),  Engr.,  Silica 
Gel  Corp.,  Baltimore  Trust  Bldg.,  and  (for  mail), 
4203  Linkwood  Rd.,  Baltimore,  Md. 

LEEK,  Walter  (M  1903),  Managing  Director  (for 
mail),  Leek  &  Co.,  Ltd.,  Htg.  Engrs.,  1111 
Homer  St.,  and  4769  W.  Second  Ave.,  Vancouver, 
B.  C.,  Canada. 

LEES,  Herbert  K.  (M  1924;  J  1912),  (for  mail), 
548  Washington  Blvd.,  and  5855  N.  Kenneth 
Ave.,  Chicago,  111. 

LEES,  John  T.  (J  1935;  5  1933),  1218  N.  New  St., 
Bethlehem,  Pa.F  and  (for  mail),  60  Boylston  St., 
Cambridge,  Mass. 

LEFFINGWELL,  Robert  R.  (J  1935;  S  1933), 
2747  Sedgwick  Ave.,  New  York,  N.  Y. 

LEGLER,  Frederick  W.  (A  1933),  Mgr.,  Retail 
Sales  (for  mail),  Waterman- Waterbury  Co.,  1121 
Jackson  St.  N.E.,  and  2919  Johnson  St.  N.E., 
Minneapolis,  Minn. 

LEIGH,  Robert  L.  (J  1935;  5  1933),  1034  Cadillac 
Dr.,  Grand  Rapids,  Mich. 

LEILICH,  Roger  L.  (M  1922),  Pres.  (for  mail), 
Baltimore  Heat  Corp.,  2000  W.  Pratt  St.,  and 
2810  Elsinor  Ave.,  Baltimore,  Md. 

LEINROTH,  J.  Paul  (M  1929),  Gen.  Industrial 
Fuel  Repr.  (for  mail),  Public  Service  Electric  & 
Gas  Co.,  SO  Park  PL,  Newark,  and  37  The 
Fairway,  Montclair,  N.  J, 

LEITCH,  Arthur  S.  (M  1908),  Pres.  and  Manag- 
ing Director  (for  mail),  The  Aithur  S.  Leitch  Co., 
Ltd.,  1123  Bay  St.,  and  421  Russell  Hill  Rd., 
Toronto,  Ont,  Canada. 

LELAND,  Warren  B.  (M  1929),  Sales  Engr.,  The 
H.  B.  Smith  Co.,  Westfield,  and  34  Leyfred 
Terrace,  and  (for  mail),  P.  O.  Box  1522,  Spring- 
tield,  Mass. 

LELAND,  William  E.  (M  1915),  Partner  (for 
mail),  Leland  &  Haley,  58  Sutler  St.,  San 
Francisco,  and  704  The  Alamecla,  Berkeley, 
Calif, 

LBNNON,  Joseph  O.  (M  1929),  Mgr.  (for  mail), 
llg  Electric  Vtg.  Co.,  15  Park  Row,  and  180 
West  f>l)th  St.,  New  York,  N.  Y, 

LEONARD,  J,  1L  (M  1931),  Mgr.  (for  mail),  J.  H, 
Leonard  Co.,  508  Scott  Bldg.,  and  844  Grosvenor 
Ave.,  Winnipeg,  Man.,  Canada. 

LEOPOLD,  Charles  S.  (M  1934),  213  S.  Broad 
St.,  Philadelphia,  Pa. 

LESLIE,  Donald  E.  (J  1935;  6'  ,1933),  3541 
Bloomington  Ave.,  Minneapolis,  Minn. 

LEUPOLD,  Herbert  W.  (J  1933),  Engr.,  Metro- 
politan Life  Insurance  Co.,  1  Madison.  Ave.,  New 
York,  and  (for  mail),  35-15-MOth  St.,  Flushing, 
N,  Y. 

LEVY*  Marion  L  (J  1931),  Salea  Mgr.  arid  C. 
Kngr.  (for  mail),  Air  Controls,  Inc.,  Div.  of 
Cleveland  Heater  CQ,,  ItWO  West  114th  St.,  and 
1273  West  108th  St.,  Cleveland,  Ohio. 

LEWIS,  Carroll  E,  (M  1930),  Pres,  (for  mail), 
Lewis  Air  Conditioners,  Inc.,  829  Second  Ave.  S., 
Minneapolis,  and  145-1  CUelmaford  St.,  St.  Paul, 
Minn. 

LEWIS,  Georfte  C,  (M  1919),  Vice-Prea.  and 
Treaa.  (for  mail),  American  Htg.  &  Vtg,  Co,, 
1505  Race  St.,  Philadelphia,  and  812  Summit 
Grove  Ave,,  Bryn  Mawr,  Pa. 

LEWIS*  Jolm  G»  (M  1026),  412  East  31st  St., 
Kansas  City,  Mo. 

LEWIS,  L.  Loftaa*  (M  1918),  Secy,  (for  mail), 
Carrier  Bngrg.  Corp*.  850  FreUnghuysen  Ave., 
Newark,  and  724  Carlton  Ave,,  Plainfkld,  N.  J, 

LEWIS,  Sarnmsl  R,*  (M  1906),  (Pruidmttot 
Mmfor)t  (Pm.»  1914;  2nd  Vlce-Pres.,  1910; 
Board  of  Governors!  1909-1910-1912;  Council, 
1914*1915),  Consulting  Engr,  (for  mail),  407 
S,  Dearborn  St.,  and  4737  JGnabark  Ave., 
Chicago,  m, 

LEWIS,    Thornton*    (M    1910),    (Presidential 
},  (Pres,,  1929;  1st  Vice-Press.,  1928;  2nd 
sn  1927;  Council,  1923-1030),  Executive 
a,  (for  mail),  Carrier  Bngrg,  Corp.,  $50 
Frellnghuynen  Ave,,  Newark,  ana  4  Halsey  PL, 
South  Orange,  N.  J. 


LIBBY,  Ralph  S.  (7  1933),  Canadian  Sumner 
Iron  Works,  Ltd.  (Stoker  Division),  560  Vernon 
Dr.,  and  (for  mail),  575  East  50th  Ave.,  Van- 
couver, B.  C.,  Canada. 

LICHTY,  Charles  P.  (M  1920),  Pres.  (for  mail). 
C.  P.  Lichty  Engrg.  Co.,  Inc.,  400^  South  21st 
St.,  and  125  Windsor  Dr.,  Birmingham,  Ala. 

LINDBERG,  Arthur  F.  (J  1935;  5  1933),  Supt., 
U.  S.  Dept.  of  Interior,  and  (for  mail),  14S4  Van 
Buren  St.,  St.  Paul,  Minn. 

LINN,  Homer  R.  (M  1914),  Engr.,  Western  Exec. 
Office,  American  Radiator  Co.,  816  S.  Michigan 
Ave.,  Chicago,  and  (for  mail),  321  S.  Ashland 
Ave.,  La  Grange,  111. 

LINTON,  John  P.  (M  1927),  Managing  Director 
(for  mail),  The  Garth  Co.,  50  Craig  St.  W.,  and 
247  Brock  Ave.  N.,  Montreal,  West,  Canada. 

LIVINGSTON,  Bernard  B.  (M  1927),  Gas  Engr. 
(for  mail),  Dept.  of  Public  Utilities,  Box  976,  and 
1630  Monument  Ave.,  Richmond,  Va. 

LLOYD,  Edward  C.  (M  1927),  (for  mail),  Arm- 
strong Cork  &  Insulation  Co.,  and  429  W. 
Walnut  St.,  Lancaster,  Pa. 

LOCKHART,  Harold  A.  (J  1935),  Engr.  (for 
mail),  Bell  £  Gossett  Co.,  3000  Wallace  St.,  and 
7906  S.  Carpenter  St.,  Chicago,  111. 

LOEFFLER,  Frank  X.  (M  1914),  Pres.  (for  mail), 
Frank  Loeffler  Supply  Co.,  710  N.  Hudson  St., 
and  320  West  26th  St.,  Oklahoma  City,  Okla. 

LOEFFLER,  Louis,  Jr.  (S  1934),  1815  W.  Ninth 
St.,  Oklahoma  City,  Okla. 

LOFTE,  John  Allen  (J  1935;  S  1933),  (for  mail), 
Carnegie  Institute  of  Technology,  Pittsburgh, 
Pa.,  and  Mondovi,  Wis. 

LOH,  Nan-Shee  (M  1933;  A  1931;  J  1927), 
House  42,  Lane  88,  Connaught  Rd.,  Shanghai, 
China. 

LONG,  David  Raymond  (M  1927),  Pres., 
Tageraft  Corp.,  142  S.  Christian  St.,  and  (for 
mail),  150  School  Lane,.  Lancaster,  Pa. 

LONGCOY,  Grant  B.  (M  1933),  Maintainance 
Engr.,  Cleveland  Board  of  Education,  and  (for 
mail),  1462  Wyandotte  Ave.,  Lakewood,  Ohio. 

LOO,  Ping  Yok  (<M  1933),  Gen.  Mgr.  (for  mail), 
China  Engrg.  Co.,  No.  35-36  Chung  Ling  Feng 
Hsin  Chia  Kow,  Chung  San  Rd.,  Nanking,  and 
271-273  Dumbarton  Rd.,  Tientsin,  China. 

LOVE,  Clarence  H.  (M  1919),  Mfrs.  Agent,,  Nash 
Engrg.  Co.,  317  Chamber  of  Commerce,  and  (for 
mail),  289  Norwalk  Ave.,  Buffalo,  N.  Y. 

LOWE,  Howard  H.  (J  1935;  S  1932),  1800  Fourth 
St.  S.E,,  Minneapolis,  Minn. 

LOWNSBERY,  Benjamin  F.  (M  1920),  Htg. 
Engr.,  Benjamin  F.  Shaw  Co.,  P.  O.  Box  953,  and 
(for  mail),  21  S.  Sycamore,  St.,  Wilmington,  Del. 

LOWY,  M.  R.  (J  1935;  3  1933),  2305  Loring  PL, 
New  York,  N.  Y. 

LUCK,  Alexander  W.*  (Life  Member;  M  1919), 
Pres.  and  Gen.  Mgr.  (for  mail),  Reading  Heater 
&  Supply  Co.,  Church  and  Woodward  Sts., 
Reading  and  Reiffton,  Pa. 

LUCKE.  Charles  E.  (M  1924),  Consulting  Engr,, 
Babcock  &  Wilcox  Co.,  85  Liberty  St.,  and 
Stevens  Prof,  of  Mech.  Engrg.  (for  mail), 
Columbia  University,  Physics  Bldg.,  and  110 
Riverside  Dr,,  New  York,  N.  Y. 

LUNJD,  Clarence  E.  (S  1933),  Lab.  Asst.,  Uni- 
versity of  Minnesota,  Experimental  Engrg. 
Bldg.,  and  (for  mail),  2729-lSth  Ave.  S.,  Minn- 
eapolis, Minn. 

LUPIENT,  Gerald  C.  (S  1935),  212  Bedford  St., 
Minneapolis,  Minn. 

LUTY,  Dpnald  J.  (M  1933) »  Asst.  Mgr.,  Air 
Cond.  Div.  (for  mail),  Gar- Wood  Industries, 
Inc.,  7924  Riopelle  St.,  Detroit,  and  911  Forest 
Ave.,  Ann  Arbor,  Mich. 

LUTZ,  James  H.,  Jr.  (M  1928),  Owner  (for  mail), 
140  Paxton  St.,  and  loOl  Forster  St.,  Harrisburg, 
Pa, 

LXJTZ,  Walter  J.  (7  1935;  5  1933),  c/o  S,  A.  E. 
House,  4915  Forbes  St.,  Pittsburgh,  Pa. 

X.  (M  1919),  Vice-Pres.,  Carrier 


XJjjG*,    J&irnQaL    J. «    v*1**    *v*'t'JL   YJ,<-C-A  *«*».,    VH»A**I*J. 

Engrg.  Corp.,  Room  408,  Chrylser  Bldg,,  New 
York,  N.  Y; 


25 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


LYLE,  J.  Irvine*  (M  1911),  (Presidential  Member] , 
(Pres.,  1917;  Council,  1917-1918),  Pres.  (for 
mail),  Carrier  Corp.,  850  Frelinghuysen  Ave., 
Newark,  and  1200  W.  Seventh  St.,  Plainfield, 
N,  J. 

LYMAN,  Samuel  E.  (.4  1924),  850  Frelmghuyscn 
Ave.,  Newark,  and  (for  mail),  728  Canton  St., 
Elizabeth,  N.  J. 

LYNCH,  William  L.  (M  1928),  Treas.  &  Genl. 
Mgr.,  (for  mail),  Rome  Turney  Radiator  Co., 
and  1413  N.  George  St.,  Rome,  N.  Y. 

LYNN,  John  H.  (A  1933),  charge  Air  Cond. 
Dept.,  (for  mail),  Texas  Automatic  Sprinkler 
Co.,  and  3431  Shenandoah  St.,  Dallas,  Texas. 

LYON,  P.  S.  (M  1929),  Commercial  Kngrg  Div., 
Air  Cond.  Dept.  (for  mail),  General  Electric  Co., 
570  Lexington  Ave.,  New  York,  and  200  Chit- 
tenden  Dr.,  Crestwood,  N.  Y. 

LYONS,  Cornelius  J.  (.1  1932),  Sales  Engr.  (for 
mail),  Nash  Engrg.  Co.,  Wilson  Ave.,  and  22 
Haviland  St.,  South  Norwalk,  Conn. 

M 

MACCUBBIN,  Howard  A.  (A/  1934),  Pres.,  H.  A. 
Maccubbin,  Inc.,  2702  Alameda  Ave.,  Baltimore, 
Md. 

MacDADE,  Ambrose  H.  (U  1923),  Sales  (for 
mail),  Burnham  Boiler  Corp.,  S.E.  Cor.  31st  and 
Jefferson  Sts.,  Philadelphia,  Pa.,  and  225  Haddon 
Ave.,  Westmont,  N.  J, 

MacDONALD,  Donald  B.  (M  1930),  C.  A,  Dun- 
ham Co.,  101  E.  Walnut  St.,  Kingston,  Pa. 

MACDONALD,  Everett  A.  (A  1933),  Br.  Mgr. 
(for  mail),  Spencer  Heater  Co.,  145  Broadway, 
Cambridge,  and  154  Standish  Rd.,  Watcrtown, 
Mass. 

MAGHEN,  James  T.  (J  1934),  Chicago  Br.  Mgr. 
(for  mail),  The  Ric-wiL  Co., 'ill  W.  Monroe  St., 
and  420  Diversey  Pkwy.,  Chicago,  111. 

MacKENZlE,  John  J.  (M  1925),  004  Shaw  St., 
Toronto,  Qnt.,  Canada. 

MacLEOD,  Kenneth  F.  (A  1033),  Mgr.,  Htg. 
Dept.,  Crane  Co.,  4 IS)  Second  Ave.  S.,  and  (for 
mail),  7703  First  Ave.  N.E.,  Seattle,  Wash. 

MAJDDUX,  Oliver  L.  (A  1TO),  Chief  En«r., 
United  Gas  £  Fuel  Co.  of  Hamilton,  Ltd.,  and 
(for  mail),  IS  Whitten  Rd.,  Hamilton,  Ont., 

MADISON,  Richard  IX  (A/  1920),  Research 
Engr.  (for  mail),  Buffalo  Forge  Co.,  490  Broad- 
way, and  133  Lisbon  Ave.,  Buffalo,  N.  Y, 

MAEHUNG,  Leon  S.  (M  1932),  Supervisor  Sales, 
Equitable  Gus  Co.,  427  Liberty  Ave.,  and  (for 
mail),  448  Sulgrave  Rtl.,  Pittsburgh,  Pa. 

MAG  INN,  Peter  F.  )Ltf«  Member;  M  1008),  Mfr«. 
Agent,  1140  8.  Neglcy  Ave.,  Pittsburgh,  )Pa. 

M^GIHL,  Willis  J.  (Af  10&4;  A  1931:  J  1027), 
Chief  Knar,  (for  mail),  P.  H.  McGJrl  Foundry  & 
Furnace  Worka,  401-1,'i  E,  Oakland  Ave.,  and  108 
Warner  Ave.,  Bloomington,  III. 

MAGNKY,  Gottlieb  R.  (M  1931),  Pres.  (for  mail), 
Ma&ney  &  Tusler,  Inc.,  Archts.  and  Kngra., 
104  S.  Ninth  St.,  and  &*2U  Waahburn  Ave,  SM 
Minneapolis,  Minn, 

MAHONEY,  David  J.  (M  1930;  A  1926),  Br, 
Mgr.  (for  mail),  Johnson  Service  Co.,  503 
Franklin  St.,  and  90  Delham  Ave.,  Buffalo,  N.  Y. 

MAIER,  George  M.  (M  1921),  AssL  to  Vice- 
Prea.  and  Gen..  Mgr.  of  Mfg.,  American  Radiator 
Co.,  8007  Jos  Campau,  Detroit,  Mich, 

MAILLARP,  Albert  L.  {M  1034).  Head  of  Air 
Cond,  Div.  (for  mail),  Kansas  City  Power  & 
Light  Co.,  1330  Baltimore,  and  3740  Washington, 
Kansas  City,  Mo. 

MAIMAN,  Herbert  (J  1935;  5  1032),  70-04-78th 
Ave,.  Glendale,  L.  I.,  N.  Y. 

MAtLIS,  William  (M  1914),  330  Lycm  BIdtf,, 
Seattle,  Wash. 

MAJLONE,  JDayte  G.  (M  1929;  A  1925),  Pres,, 
Petroleum  Heat  &  Power  Co.,  1725  S.  Michigan 
Ave*,  and  (for  mail),  7315  Merrill  Ave,,  Chicago, 

MALVIN,  Ray  C.  (M  1920),  Pres.  (for  mail), 
Mfdvirt  &  May,  Inc.,  33^  S,  Michigan  Ave,  and 
8211  Ltngley  Ave,,  Chicago,  III. 


MANAHAN,  James  E.  (.U  1934;  A  1031;  J  li>2<>), 

Air   Cond.    Engr.,    Kelvinator   Div.    (for   mail), 

Witte   Hardware   Co.,    704    N.    Third    St.,   and 

3455a  Utah  Ave.,  St.  Louis,  Mo. 
MANDEyiLLE,  Edgar  W.  (M  1914),  1171  East 

37th  St.,  Brooklyn,  N.  Y. 
MANN,  Leo  B.   (J  1030),  Air  Cond.   Engr.    (for 

mail),  Carrier  Ensrg.  Corp.,  12  South  12th  St., 

Philadelphia,    and    3018    Garrctt    Rd.,    Drcxcl 

Hill,  Pa. 
MANNING,  Walter  M.   (M  1930),   I  Its.   Kn«r., 

Salesman,  Crane  Co.,  115  K.  Front ,fcst.,  Grand 

Island,  and  (for  mail),  P.  O.  Box  112,  Clarks, 

Nebr. 
MARINO,  Dominic  A.  (J  1035;  S  1033),  755  Kast 

210th  St.,  New  York,  N.  Y. 
MARKS,  Alexander  A.  (A  1030),  Awst.  Sales  MKr. 

(for   mail),    Richmond    Radiator   Co.,    2241    N. 

American  St.,  and  GG35  McCallum  St.,   Phila- 
delphia, Pa. 
MARKUSII,  Emory  U.  (,U  1031),  Mcch  Kn«r., 

Consulting  Eni>r.  (for  mail),  225  Kast  21st  St., 

New   York,   and   S442~85th    Rd,,    Wooclhaven, 

L.  L,  N.  Y. 
MARRINER,  John  M.  S.  (H  103 n,  Sales  Kn«r., 

John  Intflis  &  Co.,  Ltd.,  14  St radian  Ave.,  and 

(for  mail)   HIM   Balsam  Ave.,   Toronto,  Out., 

Canada. 
MARSGHALL,    Peter    J.     (M    l'J30;    A     15)30; 

J   1927),    Engr.,    Krocschell    En«r«,    Co.,    23()(i 

N.   Knox  Ave,,  and  (for  mail),  2201  TouUy  Ave., 

Chicago,  111. 
MARSHALL,  IL  Hall  (M  1923),  37  West  43rd 

St.,  New  York,  N.  Y. 
MARTENIS,  John  V.  (M  1018),  Associate  Prof., 

University  of  Minnesota,  and   (for  mail),  4SOO 

Bloomington  Ave.,  Minneapolis,  Minn, 
MARTIN,  Albert  B.  (.U  11H7K  Kewaneo  Boiler 

Co.,  Inc.,  1S5S  S.  Western  Ave.,  ChU'UK»»,  111. 
MARTIN,   George  W.*   (At    11)11),   SupervimnR 

ICngr,  (for  mail),  U.  S.  Realty  ^  Improvement 

Co.,  Ill  Broadway,  New  York,  N,  V.,  ami  340 

Prospect  St.,  Ridgevvood,  N.  J. 
MARTINEZ,    Juan    J.    (J    iWitt),    P.tseo    de   la 

Reforma  183,  Mexico,  1),  F, 
MART1NKA,   Paul   I>.    (S   1U34),    13703   Cliau- 

taugua  Ave.,  Cleveland,  Ohio, 
MARTOCEtLO,    Joseph    A,    (M    11)34),    Prea., 

j.  A.   Martocello  &  Co.,  aai)  North   13th  St., 

Philadelphia,  Pa. 
MARTY,  Ed^ar  O.   (M  1016),  Ptes.  utul  Gen. 

M^r.,   Indian  Head  Anthracite,   Inc.,  and   (for 

mail),  1775  Howard  Ave.,  Pottsville,  Pa. 
MARUM,   Otto   (M   1931),    Plant    Engr,,   Agfa 

Anaco  Corp.,  21)  Oharlca  St.,  un<l  (for  nmil),  12 

Grand  Blvd.,  Bingfaamton,  N.  Y. 
MATCHETT,  James  O,  (M  1923),  Vlre-Pres.  tmd 

Gen.  IVIgr.  (for  mail),  Illinois  Engri,  ("o,,  iilst 

and  Rucinf  Ave.,  and  9930  JS.  Winchester  Avt*., 

Chicago,  HI, 
MATHER,    Harry   H,    (A    I03U),   PhiladelphiH- 

Klectric  Co.,  1000  Chestnut  St.,  Phil«dplphla,  Pa. 
MATHEY,  Nicholas  J.  (Af  U)lf»)»  Mathcy  Plbg, 

&  Htg,  Co.,  31  Third  Ave,  N,W«,  I,c  Mars,  I  own. 
MATHIS»  Euftene*  (M  1922),  New  York  Blower 

Co.,  32nd  St.  and  Shields  Av#,,  Armour  P.  O. 

Station,  Chicago,  111, 
MATHIS,    Henry    (M   1921),   The   New    York 

Blower  Co,,  32nd  and  Shields  Ave.,  and  (for 

mail),  10317  Oakley  Ave.,  Chicago,  HI. 
MATHKS,  Jfuliim  W.  (A  1021),  New  York  Blower 

Co.,  32nd  St.  and  Shields  Ave,,  Chicago.  Ill, 
MATHLS,  Victor  John  (J  1035;  S  1033),  11307  S, 

Longwood  Dr.,  Chicago,  III. 
MATTHEWS,  John  B,   (M  WB4),  DJgt,  Mgr., 

B.  F,  Sturtcvant  Co.,  1108  Commerce  Bldg,»  and 

(for  mail),  5042  Lydia  St.,  Kansas  City,  Mo, 
MAT&BISL  Harry  B*  (M  1919),  Vica-Pwi,  (for 

mail),  Carrier  BJngrg,  Corp,,  $00  Unloa  Trust 

Bldg,,  Cleveland,  and  8lft  Chadboume  Bd*. 

Shaker  Heights,  Ohio, 
MATtlttO.  Joseph  R.   (J 

North  13th  St.,  Newark,  N, 


ROLL  OF  MEMBERSHIP 


MAUER,  William  J.i:  (M  1919),  Sales  Mgr.,  Unit 

Heater  Div.  (for  mail),  C.  A.  Dunham  Co.,  450 

E.    Ohio    St.,    Chicago,    and    2525    Colfax    St., 

Evanston,  111. 
MAUTSCH,  Robert  (.1  1028),  Engr.,  Managing 

Director     (for    mail),     Compagnie     Beige    Des 

Freins    Westinghouse   97,   Avenue   Louise,  'and 

Avenue  des  Klauwaerts  38,  Brussels,  Belgium. 
MAWBY,     Pensyl     (M     1934),     Service    Engr., 

Lehigh  Navigation   Coal  Co.,   143  Liberty  St., 

New  York,  N.  Y.,  and  (for  mail),  312  Swarth- 

more  Ave.,  Ridley  Park,  Pa. 
MAXWFXL,  George  W.  (J  1935;  5  1932),  Supt. 

and  Engr.,  Kenealy  &  Maxwell,  Main  St.,  P,  O. 

Box  447,  and  (for  mail),  P.  O.  Box  422,  Harwich 

Port,  Mass. 
MAY,  Clarence  W,  (M  1938),  Pres.   (for  mail), 

May  &  Griffin,   Inc.,  501  Orpheum   Bldg.,  and 

2457  Sixth  Ave.  W.,  Seattle,  Wash. 
MAY,  Edward  M.  (M  1931),  Combustioneer,  Inc., 

1835  S.  Michigan  Ave.,  Chicago,  and  (for  mail), 

1022  N.  Hayes  Ave.,  Oak  Park,  111. 
MAY,  George  Elmer  (M  1933),  Air  Cond,  Engr. 

(for   mail),    New  Orleans   Public  Service,    Inc., 

317    Buronne    St..    and    2031    Short    St.,    New 

Orleans,  La. 
MAY,   Maxwell   F.    (M  1929),   Secy-Treas.   (for 

mail),   Malvin  &  May,   Inc.,  332  S.   Michigan 

Ave.,  Chicago,  and  Palos  Park,  111. 
MAYETTK,  Charles  E,  (A/  1926),  1400  Floral  St. 

Washington,  D.  C. 
MAYNARD,  Herbert  R.  (5  1935),  1725  Wood- 

land Ave.,  Duluth,  and  (for  mail),  1014  Seventh 

St.  S.K.,  Minneapolis,  Minn. 
MAYNARD,    J.    Earle    (M    1931),    Chief    Htg. 

Knur,,  Fox  Furnace  Co.,  and  (for  mail),  Tele- 

graph Rtl.,  Elyria,  Ohio. 
McCAULEY,   James   II.    (M  1921),   James  H. 

McCaulcy,  Inc.,  5321  West  05th  St.,  Chicago,  III. 
McCLKLLAN,  James  E.  (M  1922),  Mgr.,  Chicago 

Dist.  (for  mail),  American  Blower  Corp.,  228  N, 

LaSalle  St.,  Chicago,  and  88M  LaCrosse  Ave., 

Nilea  Center,  111. 
McCLINTOCK,  Alexander,  Jr.  (M  1928;  J  1920), 

Member  of  Firm  (for  mail),  A.  McClmtock  & 

Sons,  11)37  Ridge  Ave.,  and  121  Rochclle  Ave., 

Wiaaahickon,  Philadelphia,  Pa. 
McOLOUGlIAN,  Charles  (S  1934),  279  Ryeraon 

St.,  Brooklyn,  N,  Y. 
McOOLL,    Jay    R,*     (M    1910),     (Presidential 

Mmbf.r'),  (Pros.,  1922;  1st  Vice-Pros.,  1021;  2nd 

Vicc-Prt'8,,    11)20;    Council,    1920-1023),    2304 

Penobscot  Bldg,,  Detroit,  Mich. 
McCONACUIB,  Lorn*  L.   (A   1028)  >   Htg.  and 

Plbg.,  8817  Mack  Ave.,  ami   (for  mail),  1379 

Maryland  Ave.»  Detroit,  Mich. 
McCQNNER,  Charles  R.  (A  1925;  J  1922),  Gen. 

Salea  Mgr.  (for  mail),  Clanigc  Fan  Co.,  and  1904 

Waite  Ave.,  Kalamazoo,  Mich, 
McGORMAGK,    Denis    (M    1933),    Mgr.,    Air 

Cond.  Instruments  and  Control  Dept.  (for  mail), 

Jullen  P.  Kriez  &  Sons,  Inc.,  4  N.  Central  Ave., 

and  5924  Hellona  Ava.,  Baltimore,  Md. 
McGOY,  Thomas  F,  (M  1924),  Mgr.  (for  mail), 

The  Powers  Regulator  Co.,  125  at,  Botolph  St., 

Boston,  and  Glen  Rd,,  Wellealey  Farms,  Mass. 
McGREERY,   Hufth  J,  (U   1022),   (for  mail), 

Marine  Bldg.,  and  1617-49tli  Ave,  W,,  Van- 

couver, B«  C,»  Canada. 
McGtJNE,  Byron,  V.  (M  1928),  Sties  Engr.  (for 

mall),  101  W,  Yaklma  Ave.,  P.  O,  Box  385,  and 

2810  W.  Yaktmt  Ave,,  Yakima,  Wash. 
McDONALP,  Thomas  (A  1031),  Mgr,  (for  mail), 

Minneapolis-Honeywell    Regulator    Co,,    Ltd., 

H7  Peter  St.,  and  $(J  ICingsway,  Toronto,  Out., 

Canada* 

HeJ>QtoBIX»  Everett  N,  (M  1023),  Prea.  (for 
""  (  McDonnell  IE  Miller,  400  R  Michigan 

and  105  E,  Delaware,  Chicago,  111.  ^ 
3IN,  Joto  W*  (7  1981),  180  Rowland  Park, 

F.  (A  1928),  1258  Pratt 


Mi,  Smith  ft  MoOinn^e  Co.,  527  JWwt 
n  and  142  Btllleld  Ave*,  Pittsburgh,  Pa, 


McGLENN,  G.  Raymond  (M  1915),  259  Lormore 
St.,  Elmira,  N.  Y. 

McGONAGLE,  Arthur  (M  1932),  Consulting 
Engr.  (for  mail),  1013  Fulton  Bldg.,  Pittsburgh, 
and  6815  Prospect  Ave.,  Ben  Avon,  Pa. 

McGRAIL,  Thomas  E.  (M  1926),  Mgr.  Htg. 
Dept.,  Crane,  Ltd.,  Beaver  Hall,  and  (for  mail), 
W.  A.  90S7-34G5  Belmore  Ave.,  Montreal,  P.  Q., 
Canada.  " 

McGUIGAN,  L.  A.  (A  1919),  724  Hastings  St., 
Pittsburgh,  Pa. 

McHENRY,  Robert  W.  (M  1921),  Engr.,  Trans. 
Canada  Radiator  &  Boiler  Co.,  672  Dupont  St., 
and  (for  mail),  236  Eglinton  Ave.,  Toronto,  Ont., 
Canada. 

MclLVAINE,  John  H.*  (A/  1029),  Prcs.  (for  mail), 
McIIvaine  Burner  Corp.,  OG3  W.  Washington 
Blvd.,  and  1100  Lake  Shore  Dr.,  Chicago,  111. 

MclNTIRE,  James  F.  (M  1915;  A  1914),  (Coun- 
cil, 192G-192S,  1932-1934),  Vice-Pres.  (for  mail), 
U.  S.  Radiator  Corp.,  1056-44  Cadillac  Square, 
P.  O.  Box  686,  and  3261  Sherbourne  Rd.,  Detroit, 
Mich. 

McINTOSH,  Fabian  C.  (Al  1921;  J  1917), 
(Treas.,  1930;  Council,  1929-1934),  Br.  Mgr.  (for 
mail),  Johnson  Service  Co.,  1238  Brighton  Rd,, 
and  302  Marshall  Ave.,  Pittsburgh,  Pa. 

McKIEVER,  William  H.*  (M  1897;  J  1896), 
Pres.  (for  mail),  William  H.  McKiever,  Inc.,  247 
West  13th  St.,  New  York,  and  479  Eighth  St., 
Brooklyn,  N.  Y, 

McKINLEY,  C.  B.  (5  1934),  132  Paye,  Norman, 
Okla. 

McKINNEY,  William  J.  (.1  1934),  Mgr.,  Atlantic 
Dist.,  American  Blower  Corp.,  710-101  Marietta 
St.  Bldg.,  Atlanta,  Ga. 

McKOTRICK,  Percy  A.  (A  1934),  Treas.  and 
Gen.  Mgr.  (for  mail),  Parks-Cramer  Co.,  970 
Main  St.,  and  219  Blossom  St.,  Fitchburg,  Mass. 

McLARNEY,  Harry  W.  (M  1933),  Air  Cond. 
Engr.  (for  mail),  Union  Electric  Light  &  Power 
Co.,  315  North  12th  St.,  and  5053  Lindcnwood 
Ave.,  St.  Louis,  Mo. 

MCLAUGHLIN,  Joseph  D.  (A  1930;  J  loss), 

Owner  (for  mail),  Bralcy  &  McLaughlin,   166 

Aborn  St.,  and  45  Roslyn  Ave.»  Providence,  R.  I, 
McLEAN,  Dcrmld  (M  1917),  McColl,  Snyder  c% 

McLean,  2304  Penobscot  Bldg.,  Detroit,  Mich. 
McLEISH,  William  S.  (A  1932:  J  1928),  Dist. 

Engr.  (for  mail),  The  Ric-wiL  Co.,  Room  1838, 

101   Park  Ave.,  and  500  Riverside  Dr.,  Ne'w 

York,  N.  Y. 
McLENEGAN,  David  W.  (M  1933),  Asst.  Engr., 

Air  Cond.  Dept.  (for  mail),  General  Electric  Co., 

1  River  Rd.,  and  16SG  Rugby  Rd.,  Schenectady, 

N.  Y. 
McLOUTH,    Bruce    F.    (J    1934),    Dail    Steel 

Products  Co.,  Lansing,  Mich. 
McMAHON,  Thomas  W.  (M  1928),  Dist,  Mgr. 

(for  mail),  American  Blower  Corp.,  1715  Railway 

Exchange   Bldg.,   and   0151   Waterman  Ave., 

St.  Louis,  Mo, 
McMUNN,   A.   H,    (S   1934),    (for   mail),   4915 

Forbes  St.,  Pittsburgh,  Pa.,  and  311  St.  Clair  St., 

Cl&rkuburg,  W.  Va. 
McMURRER,  Louis  J.  (M  1928:  A  1928 ;  J  1924), 

Pres.,  The  McMurrer  Co,,  303  Congress  St,, 

Boston,  and   (for  mail),  190  Harvard  Circle, 

Newtonvllle,  Mass. 
McNAMARA,  William  (A  1030),  Sales  Engr.  (for 

mail),  The  Trane  Co,,  2694  University  Ave.,  and 

1855  Como  Ave*  W,,  St.  Paul,  Minn. 
McPHERSON,   William   A.    (M   1929),  Chief, 

Htg,  and  Vtg.  Div.,  Dept,  of  School  Bldgs.,  H 

Beacon  St.,  Boston,  and  (for  mail),  86  Dwinnell 

St»  West  Roxbury,  Mass. 
McOUAlD,  Dan  J.   (M  1934),  Engrg.  Service 

Work,  1505  Milwaukee  St.,  Denver,  Colo, 
McTERNAN,  Felix  J,  (A  1981),  1523  Main  St., 

Buffalo,  N,  Y, 
MKADt  Bd<ward  A.  (M  1$26>,  Asst  Sales  Mgr.  (for 

mail),  Nash  Engrg.  Co.,  and  5  Thames  St., 

Norwalk,  Conn. 


27 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


MEFFERT,  George  H.  (7  1930),  Engr.  (for  mail), 

Carrier  Engrg.  Corp.,  2022  Bryan  St.,  and  41543^ 

Prescott  Ave.,  Dallas,  Texas. 
MEHNE,    Carl    A.    (M    1929),    Htg.    and    Vtg. 

Expert  (for  mail),  Room  821,   101  Park  Ave., 

New  York,  and  Livingston  St.,  Valhalla,  N.  Y. 
MEINKE,   Howard  G.    (M  1933),  Asst.   Engr., 

Civil  Engrg.  Dept.  (for  mail) ,  New  York  Edison 

Co.,  4  Irving  PL,  Room  1517-S,  New  York,  and 

41  Harte  St.,  Baldwin,  Nassau  Co.,  N.  Y. 
MEISEL,  Carl  L.  (J  1931),  350  Central  Park  West, 

New  York,  N.  Y. 
MELLON,    James   T.   J.    (M    1911),    (Council, 

1915),  (for  mail),  Mellon  Co.,  4415-21  Ludlow 

St.,  and  431  North  63rd  St.,  Philadelphia,  Pa. 
MENSING,    Frederick   D.    (M   1920),    (Treas., 

1931-1932;  Council,  1931-1932),  Consulting  Engr. 

(for  mail),   Mensing  &  Co.,  12  South  12th  St., 

and  2845  Frankford  Ave.,  Philadelphia,  Pa. 
MERKEL,   Fred  P.    (M  1924),  204-llth  Ave., 

Belmar,  N.  J. 
MERLE,  AndrS   (M  1934),  Director  of  Engrg., 

Control  Corp.  ot  America,  250  West  57th  St., 

New  York,  N.  Y.,  and  (for  mail),  172  Lincoln 

Ave.,  Elizabeth,  N.  T. 
MERRILL,  Carle  J.  (M  1919),  Treas.  (for  mail), 

C.   J.   Merrill,    Inc.,   54   St.   John  St.,  and   15 

Longfellow  St.,  Portland,  Maine. 
MERRILL,  Frank  A.  (M  1934),  Consulting  Htg. 

and  Vtg.  Engr.  (for  mail),  Hollis  French,  210 

South   St.,    Boston,   and   19   Auburndale   Rd., 

Marblehead,  Mass. 
MERRITT,  G,  J.   (M  1925),  Director,  Merritt, 

Ltd.,  8  French  Bund,  Shanghai,  China. 
MERTZ,  Walter  A.  (M  1919),  Kehm  Bros,  51  E. 

Grand  Ave.,  Chicago,  111. 
MERWIN,    Gile   E.    (M    1924;   J   1923),    Secy., 

Rockford    Plbg.    Supply    Co.,    and    (for   mail), 

1325~20th  St.  Rockforcl,  111. 
MEYER,  Charles  L.  (M   1930),   198-25   Foothill 

Terrace,  Hollis,  L.  I.,  N.  Y. 
MEYER,  Frank  L.  (M  1932;  J  1928),  Viee-Pres.. 

The  Meyer  Furnace  Co,,  and  (for  mail),  i)  Cole 

Court,  Peoria,  111. 
MEYER,   Henry  C.,   Jr.*   (M   1898),   (Council, 

1915-1916),  Pres.   (for  mail),  Meyer  Strong  & 

Jones,  Inc.,  101  Park  Ave.,  New  York,  N.  Y.,  and 

25  Highland  Ave.,  Montclair,  N.  J. 
MEYER,    John    W.    (A    1929),    Asst.    to    Mgr., 

Industrial  Sales  Dcpt.  (for  mail),  Philadelphia 

Electric  Co.,  1000  Chestnut  St.,  and  5000  Pine 

St.,  Philadelphia,  Pa. 
MICHIE,  IX  Fraser  (A  1930),  Boiler  and  Rad. 

Div.,  Crane,  Ltd.,  &J  Lombard  St.,  and  (for 

mail),  5  B,  553  Wardlaw  Ave.,  Winnipeg,  Man,, 

Canada. 
MILES,  James  C.   (M  1914),  Dept.  Mgr.   (for 

mail),    The   Henry   Furnace   &   Foundry   Co., 

3471  East  49th  St.,  and  1803  Crawford   Rd., 

Cleveland,  Ohio, 
MILLAR,  Rowland  J.  (M  1925),  Mgr.  (for  mail). 

Pease  Foundry  Co,,  Ltd.,  118  King  St.  K,,  and 

S,?Oakmount  Rd,  Toronto,  Ont,,  Canada. 
MILLARD,  Jimius  W,  (HI  1929),  Diet.  M«r.  (for 

mail),  Carrier  Engrg.  Corp.,  410  Asylum  St., 

Hartford,  and  Manchester,  Conn. 
MILLER,  Bruce  R.  (A  1930),  1533  N.W.  25th  St., 

Oklahoma  City,  Okla, 
MILLER,  Charles  A.   (A  I«17),  Salesman  (for 

mail).  The  H.  B.  Smith  Co.,  10  East  41st  St.,  and 

2870  Marion  Ave.,  New  York,  N.  Y, 
MILLER,  Charles  W.   (M  1919;  J  1908),  (for 

mail),  The  Rado  Co.,  388  S.  Second  St.,  Mil- 
waukee, and  R,  1,  Box  42,  Menomonee  Falls, 

Wta. 
MILLER,  Floyd  A,  (M  1011),  477  Federal  Bldg., 

Chicago,  111. 
MILLER,  Harold  A,  (S  1935),  1000-24th  Ave, 

S,E,»  Minneapolis,  Minn. 
MILLER,  Harry  M,  (M  1920),  3938  N.  Stowell 

Ave.,  Milwaukee,  Wis. 
MILLER,  James  E.  (M  1914;  J  1912),  Vice-Prea, 

(for  mall),  C.  W,  Johnson,  Inc.,  2>11  N.  Des- 

plainea   St.,    Chicago,   and    2210   Colfax   St., 

Evanston,  III, 


MILLER,  John  F.  G.  (M  1916),  Vice-Prcs.  (for 

mail),  B.  F.  Sturtevant  Co.,  Hyde  Park,  Boston, 

and  20  Chapel  St.,  Brookline,  Mass. 
MILLER,  Leo  B.  (M  1926),  Refrigeration  Div.  (for 

mail),    Minneapolis-Honeywell    Regulator    Co., 

2753  Fourth  Aye.  S.,  and  2010  James  Ave,  S., 

Minneapolis,  Minn. 
MILLER,  Lorin  G.  (M  1933),  Prof.  Mech.  Engrg. 

(for  mail),   Dept.   of   Mech.   Engrg.,    Michigan 

State   College,    Engrg.    Bldg.,   and   920    Sunset 

Lane,  E.  Lansing,  Mich. 
MILLER,  Merl  W.   (M  1932;  J  1920),  Mgr.  of 

Lab.  (for  mail),  Trane  Co.,  and  229  South  St., 

LaCrosse,  Wis. 
MILLER,  Robert  A.*  (M  1931),  Tech.  Sales  Engr. 

(for   mail),    Pittsburgh   Plate   Glass    Co.,    2200 

Grant  Bldg.,  Pittsburgh,  and  1211  Carlisle  St., 

Tarentum,  Pa. 
MILLER,  Robert  T.  (A  1927),  Chief  Engr.  (for 

mail),  Masonite  Corp.,  Ill  W.  Washington  St., 

Chicago,    and    1228    Sunnyside   Ave.,    Chicago 

Heights,  111. 
MILLER,  Tolbert  G.   (A  1929;  J  1921),  Supt., 

Htg.  and  Vtg.,  11  N.  Second  St.,  Wormleysburg, 

Pa. 
MILLIKEN,  James  H.*  (M  1923),  Dist.  Repr. 

(for  mail),  American  Air  Filter  Co.,  Inc.,  20  N. 

Wacker  Drive,  Chicago,  and  1021   Ridge  Ct., 

Evanston,  111. 
MILLIKEN,  Vincent  D.  (A  1930),  Sales  Mgr.  (for 

mail),  Skidmore  Corp.,  and  2015  Forres  Ave., 

St.  Joseph,  Mich. 
MILLIS,  Lirm  W.  (Life  Member;  M  1018),  Secy., 

Security  Stove  &  Mfg.  Co.,  1U30  Oakland,  and 

(for  mail),  3534  Wabaah  Ave,,  Kansas  City,  Mo, 
MILWARD,  Robert  K.  (A  1920),  Mgr.  (for  mail), 

U.  S.  Radiator  Corp.,  127  Campbell  Ave.,  and 

2441  Culvert  Ave.,  Detroit.  Mich. 
MINER,  Major  H.  (,S  1934),  1510^  North-wuat  25, 

Oklahoma  City,  Okla. 
MITCHELL,  C.  II.  (A/  1924),  Engr.,  The  Fels  Co., 

42  Union  St.,  Portland,  Maine,  and  (for  mail), 

179  Thatcher  St.,  Milton,  Mass. 
MITTENDORFF,  E.  M.  (M  1932),  Sules  Engr., 

ffor  mail),  Sarco  Co.,  Inc.,  222  N.  Bank  Dr., 

Chicago,  and  1000  S.  Prospect  Ave.,  Park  Ridge, 

III.      ' 
MJOLSNES,    Leonard   O.    (S   1035),    018  15th 

Ave.  S.K.,  Minneapolis,  Minn. 
MOOIANO,  Rene  (M  1925),  55  Boulevard  Beau- 

sejour,  Paris,  10,  erne,  France. 
MOLER,   William  H,    (Af   1927;   J   1923),    Bn 

Supervisor    Com.    Div.,    York    Ice    Machinery 

Corp.,  2225  S.  Lamar  St.,  Dallas,  ami  (for  mail), 

R.  K,  D.  1,  Box  37B,  Irving,  Texas. 
MONDAY,    Charles    E.    (M    1920),    Chas,    E, 

Monday  &  Co.,  1328  Fairmount  Ave.,  Phila- 
delphia, Pa. 
MONROE,  Meade  (J  1035;  S 1933),  1228  Southern 

Blvd.,  New  York,  N.  Y. 
MONROE,  Raymond  R,  (A  1929),  7  County  St., 

Ipawich,  Mass. 
MONTGOMERY,  Ora  C,  (M  1933),  Asst.  Supt. 

of  Power  (for  mail),  N.  Y.  C,  R,  R.,  Grand 

Central  Terminal,  Room  1$4#»  and  255  West 

84th  St.,  New  York,  N,  Y, 
MOODY,  Lawrence  E,   (M  1919),  Member  of 

Firm,    Isaac    Hathaway    Francis,    Consulting 

EnKra.,  Otis  Bldg.,  Philadelphia,  Pa,»  and  (for 

mail),  237  Jefferson  Ave.,  Haddonfteld.  N.  J. 
MOON,  L.  Walter  (M  1915),   (Council,   1933- 

1034),  Pres,  (for  mail),   Bradley  Heating  Co., 

3884  Olive  St.,  and   5006  N.   Kingshigkwty, 

St.  Louis,  Mo. 
MOORE,  Henry  W*  (M  1935),  Mr  Cond,  Engr. 

(for  mail),  Frigldalre  Corp»,  Dayton,  Ohio,  and 

«16  Greenland  Dr.,  Murfre^sboro,  Term. 
MOORE,  Herbert  S.  (A  I923)»  Mfr».  Agent,  107 

Clendenan  Ave.,  Toronto  0>  Ont.,  Canada* 
MOORE,  Robert  E*  (J  1933),  Junior  Sale*  Bn«r, 

(Div,   of),   Manning  Mwcweu   &   Moore,   446 

Communlpaw  Ave,,  Jersey  City,  N.  J,»  a»d  (for 

mail),  1730  Bast  46th  St.,  Brooklyn,  N.  Y» 
MOORB,  Robert  B.  (A  1928),  714  Brumwel  SU 

Evanston,  111. 


28 


ROLL  OF  MEMBERSHIP 


MOREAU,  Donate  (-4  1932),  35  E.  McCormick, 
Tucson,  Ariz. 

MORGAN,  Glenn  C.  (M  1911),  Partner  (for  mail), 
Morgan-Gerrish  Co.,  307  Essex  Bldg.,  and  4308 
Fremont  Ave.  S.,  Minneapolis,  Minn! 

MORGAN,  Robert  C.  (M  1915),  314  W.  Seymour 
St.,  Philadelphia,  Pa. 

MOREHOUSE,  H.  Preston  (M  1933),  General 
Air  Cond.  Repr.  (for  mail),  Public  Service  Elec. 
&  Gas  Co.,  80  Park  PL,  Newark,  and  S5  Halsted 
St.,  East  Orange,  N.  J. 

MORRIS,  Arnold  M.  (J  1934),  Sheet  Metal 
Worker,  Philadelphia  Navy  Yard,  Sheet  Metal 
Shop  Building  No.  17,  and  (for  mail),  3022  Baltz 
St.,  Philadelphia,  Pa. 

MORRIS,  Edward  J.  (7  1935;  5  1931),  Engr., 
Morris  Engrg.  Co.,  Inc.,  107  E.  Pleasant,  and 
(for  mail),  3414  Gwynn's  Falls  Pkwy.,  Balti- 
more, Md. 

MORRIS,  Fred  H.  (A  1929),  14704  Stratmore 
Ave.,  E.  Cleveland,  Ohio. 

MORRISON,  Chester  B.  (M  1931),  Mgr.  (for 
mail),  York  Shipley  Co.,  81  Jinkee  Rd.,  and 
347  Route  Cohen,  Shanghai,  China. 

MORSE,  Clark  T.  (M  1913),  Pres.  (for  mail), 
American  Blower  Corp.,  6000  Russell,  and 
16222  Shaftsbury  Rd.,  Detroit,  Mich. 

MORSE,  Floyd  W.  (A  1934),  (for  mail),  Chamber- 
lin  Metal  Weather  Strip  Co.,  52  Vanderbilt  Ave., 
New  York,  and  112  Sycamore  Ave.,  Mt.  Vernon, 
N.  Y. 

MORTON,  Charles  H.  (A  1931),  HOG  Sherman 
St,  S.E.,  Grand  Rapids,  Mich. 

MORTON,  Harold  S.  (M  1931),  Dist.  'Mgr., 
Modern  Coal  Burner  Co.,  538  Baker  Bldg.,  and 
(for  mail),  4330  Wooddale  Ave.,  Minneapolis, 
Minn, 

MOSHBR,  Clarence  H.  (A  1919),  C.  H.  Mosher 
Co.,  423  Ashland  Ave.,  Buffalo,  N.  Y. 

MOSS,  Edward  (M  1920),  1130  Atlantic  Ave., 
Brooklyn,  N.  Y. 

MOTZ,  O.  Wayne  (M  1932),  Mech.  Engr,, 
Samuel  Hannuford  &  Sons,  Archts-,  1024  Dixie 
Terminal  Bldg.,  Cincinnati,  and  (for  mail),  2587 
Irving  PL,  Norwood,  Ohio. 

MOULDER,  Albert  W.*  (M  1917),  Mgr.,  Htg., 
Power  and  Industrial  Piping  Div.  (for  mail), 
Grinnell  Co.,  Inc.,  260  W.  Exchange  St.,  and  12 
Blackstone  Blvd.,  Providence,  R,  L 

MOULTON,  David  (M  1920),  99  Chauncy  St., 
Boston,  Masa. 

MUELLER,  Harold  C*  (A  1930),  Sales  Engr.  (for 
mail),  Powers  Regulator  Co.,  2720  Greenview 
Ave»»  Chicago,  and  2720  Lawndale  Ave,,  Evans- 
ton,  IU. 

MUNDER,  John  F.,  Jr.  (M  1927;  J  1024),  (for 
mail),  Quinn  Engrg,  Co.,  501  Madron  Ave., 
New  York,  N,  Y.,  and  81  Joyce  Rd,,  Tenafly, 

MUNIBtL  Leon  L.  (M  1910;  J  1916),  Pres,  (for 
mall),  Wolff  &  Munier,  Inc..  222  East  41st  St., 
New  York,  and  63  Columbia  Ave,,  Hartsdale, 
N.  Y, 

MUNEO,  Edward  A.  (Charter  Member;  Life 
Member},  Htg,  and  Vtg.  Engr,,  c/o  Arthur  B. 
Munro,  50  Jarvls  PL,  Lynbrook,  L,  L,  N.  Y. 

MURPHY,  Charles  G.  (S  1984),  3415  Fort 
Independence  St.,  New  York,  N*  Y  . 

M0RPHY,  Edward  T**  (M  1915),  Vice-Pres.  (for 
mall).  Carrier  Engrg,  Corp.,  180  N,  Michigan 
Ave,,  and  200  E.  Chestnut  St.,  Chicago,  111, 

M0RPHY,  Howard  0,*  <M  1928),  Vice-Prea.  (for 
mail),  American  Air  Filter  Co.,  me*,  215  Central 
Ave*,  and  4W  Ughtfoot  Rd,,  Louisville,  Ky. 

MOTJPHY,  Joaepfe  R.  (M  1034J  A  1925),  The 
Terrace,  Riverside,  Conn. 

MtWHY,  William  W,  (M  1980),  Trees,  (for 
mall).  W.  W.  Murohy  Co*,  171  Chestnut  St.,  and 
m  Mansfield  St.,  Springfield,  Ma«u 

MURRAY*  John  J,  (A  1938),  Salesman,  Vice- 
Pjres>t  Pierre  Perry  Co,,  236  Congress  St*,  Boston, 
and  (for  mail),  00  Commonwealth  Park  W., 
Newton  Center,  Mass, 


MURRAY,  Thomas  F.  (M  1923),  State  Architect, 

14  S.  Lake  Ave.,  Albany,  N.  Y. 
MUSGRAVE,   Merrill  N.    (A    1935),    Pres.    (for 

mail),  Harrison  Sales  Co.,  314  Ninth  Ave.  N.,  and 

140  East  64th  St.,  Seattle,  Wash. 
MYERS,  Frank  L.  (M  1933),  Sales  Engr.,  Owens, 

Illinois  Glass  Co.,  and  (for  mail),  3406  Detroit 

Ave.,  Toledo,  Ohio. 
MYERS,  Charles  R.  (S  1935),  White  Bear  Lake, 

R.  No.  2,  Minneapolis,  Minn. 
MYERS,  George  W.  F.  (M  1930;  A  1928;  J  1923), 

Mfrs.  Repr.,  Htg.,  Vtg.,  and  Air  Cond.,  Mart 

Bldg.,  401  South  12th  St.,  St.  Louis,  and  (for 

mail),  476  Pasadena  Ave.,  Webster  Groves,  Mo. 

N 

NAROWETZ,  Louis  L.,  Jr.  (M  1929;  A  1912), 
Secy,  (for  mail),  Narowetz  Htg  &  Vtg.  Co., 
1711  Maypole  Ave.,  Chicago,  and  112  S.  Park 
Ave.,  Park  Ridge,  111. 

NASON,  George  L.  (M  1929;  A  1929;  J  1927),  31 
N.  Franklin  St.,  Holbrook,  Mass. 

NASS,  Arthur  F.  (M  1927),  Secy-Treas.  (for 
mail),  McGinness,  Smith  &  McGinness  Co.,  527 
First  Ave.,  Pittsburgh,  and  Elmhurst  Rd., 
R.  P.  No.  8,  Crafton  P.  O.,  Pa. 

NATKIN,  Benjamin*  (M  1909;  J  1907),  Pres.  (for 
mail),  Natkin  &  Co*,  2020  Wyandotte,  and  5211 
Rockhill  Rd..  Kansas  City,  Mo. 

NAYLOR,  Charles  L.  (M  1931),  Supt.,  Heat, 
Light  and  Power  (for  mail),  The  Atlantic 
Refining  Co.,  3144  Passyunk  Ave.,  and  2315 
North  ISth  St.,  Philadelphia,  Pa. 

NEALE,  Laurence  I.  (A  1927),  Vice-Pres,  (for 
mail),  Atlantic  Gypsum  Products  Co.,  60  East 
42nd  St.,  and  125  East  57th  St.,  New  York,  N.  Y. 

NEARY,  Daniel  A.  (J  1935;  5  1933),  444  East 
66th  St.,  New  York,  N.  Y. 

NEEDLER,  J.  H.  (M  1933),  Phillips  Getschow 
Co.,  32  W.  Austin  Ave.,  Chicago,  111. 

NEILER,  Samuel  G.  (M  1898).  Consulting  Mech. 
and  Elec.  Engr.  (for  mail),  Neiler,  Rich  &  Co., 
431  S.  Dearborn  St.,  Chicago,  and  737  N.  Oak 
Park  Ave.,  Oak  Park,  111. 

NELSON,  Chester  L.  (/  1929),  6704  Oconto 
Ave.  N.,  Chicago,  111. 

NELSON,  D.  W,*  (M  1928),  Asst.  Prof,  of  Steam 
and  Gas  Engrg.  (for  mail),  Mech.  Engrg.  Bldg., 
University  of  Wisconsin,  and  3906  Council 
Crest,  Madiaon,  Wis. 

NELSON,  George  O.  (M  1923),  Carstens  Bros., 
Ackley,  Iowa. 

NELSON,  Harold  A.  (M  1926),  236  S.  La  Pere  St., 
Beverly  Hills,  Calif. 

NELSON,  Herman  W.  (M  1909),  Pres.  and  Gen 
Mgr.  (for  mall),  The  Herman  Nelson  Corp.,  1824 
Third  Ave,,  and  2500-llth  St.,  Moline,  111. 

NELSON,  Raymond  Allen  (S  1935),  14th  St.  and 
Prospect  Ave,,  Cloquet,  ana  (for  mail),  418-18th 
Ave.  S.E.,  Minneapolis,  Minn. 

NELSON,  Richard  H,  (A  1033;  J  1928),  Secy- 
Treas,.  Herman  Nelson  Corp.,  1824  Third  Ave., 
and  (for  mail),  1303«30th  St.,  Moline,  111. 

NESBITTj  Albert  J,*  (M  1921;  J  1921),  Secy- 
Treas,  (for  mail),  John  J.  Nesbitt,  Inc.,  State 
Rd,  and  Rhawn  St.,  and  Matchwood  Apts*, 
Wissahickon  Ave.,  and  School  Lane,  Phila- 
delphia, Pa, 

NESBITTf  John  J.  (M  1923),  John  J,  Nesbitt, 
Inc.,  State  Rd.  and  Rhawn  St.,  Holmesburg, 
Philadelphia,  Pa. 

NESDAHL,  Eilett  (M  1915),  c/o  Colben  Nesdahl, 
Route  it  Shevlin,  Minn4 

NBSS,  William  H.  G,  (M  1931),  Gen.  Mgr,  (for 
mail),  Master  Fan  Corp.,  1323  Channing  St.,  and 
215  N,  Kingaley  Dr,,  Los  Angeles,  Calif. 

NESSI,  A*idr6  (M  1930),  Ingr  des  Arts  et  Manu- 
factures, Expert  prea.  le  triblnal  civil  de  la  Seine, 
and  (for  mail),  1  Avenue  du  President  Wilson, 
Paris  XVI,  France, 

NBU,  Henri  J.  E.  (M  1933),  Pres,,  Etablissements 
Neu,  4.7-49  Rue  Fourier,  Lille  (Nord),  France. 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


NEWCOMB,  Lionel  B.  (J  1933),  Junior  Engr., 

Philadelphia  Electric  Co.,  and  (for  mail),  6056 

Walton  Ave.,  Philadelphia,  Pa. 
NEWPORT,  Charles  F.*  (U  1900),  Sales  Engr., 

Weil-McLain  Co.,  Michigan  City,  Ind.,  and  (for 

mail),  ,10001  Longwood  Dr.,  Chicago,  111. 
NICELY,  John  E.   (A   1925),   120S  Marion  St., 

Reading,  Pa. 
NICHOLLS,  Percy*  (M  1920),  Supervising  Engr., 

Fuel  Section  (for  mail),  U.  S.  Bureau  of  Mines, 

Pittsburgh,  Pa. 
NIGHTINGALE,  George  F.  (A  1931),  Western 

Sales  Mgr.,  Tuttle  &  Bailey,  Inc.,  61  W.  Kwizie 

St.,  Chicago,  and  (for  mail),  621  S.  Maple  Ave., 

Oak  Park,  111. 
NOBBS,    Walter    W.    (M    1919),    50    Fairhazcl 

Gardens,  London  N.W.6,  England. 
NOBIS,  Harry  M.  (M  1914),  1827  Stanwood  Rd., 

East  Cleveland,  Ohio. 
NOBLE,  Theodore  G.  (J  1935;  S  1933),  Ensr., 

Minneapolis  Gas  Light  Co.,  and  (for  mail),  523 

Oak  St.  S.E.,  Minneapolis,  Minn. 
NOLL,  William  F.  (M  1924),  Htg.  Contractor, 

2850  North  47th  St.,  Milwaukee,  Wis. 
NORDHEIMER,  Clyde  L.  (/  1935;  S  1931),  622 

Mellon  St.,  Pittsburgh,  Pa. 
NORRIS,   William  D,    (M   1930),    1314   Forest 

Ave.,  Wilraette,  111. 
NORTHON,  Louis  (AT  1929),  Consulting  Engr., 

132  Park  Ave.,  Mt.  Vernon,  N.  V. 
NOTTBERG,  Gustav  (A  1033),  Secy,  (for  mail), 

U.  S.  Engineering  Co.,  914  Campbell,  and  1835 

East  68th  St.  Terrace,  Kansas  City,  Mo. 
NOTTBERG,  Henry  (M  1919),  Vice-Pres.   (for 

mail),  U.  S,  Engineering  Co.,  014  Campbell  St., 

and  213  South  Bales,  Kansas  City,  Mo. 
NOVOTNEY,    Thomas    A.    (M    1928),    Mgr., 

Research  and    Sales   Engrg,   Depts.,   National 

Radiator  Corp.,  221  Central  Ave.,  and  (far  mail), 

403  Wayne  St.,  Johnstown,  Pa. 
NOWITSKY,  Herman  S.  (A  1931),  Supt.,  Con- 
struction, Repairs  and  Maintenance,  Wilmers  & 

Vincent  Corp.,  and   (for  mail),  151  Tenth  St., 

Norfolk,  Va. 
NUSBAUM,  Lee*  (M  1915),  Owner  (for  mail), 

Pennsylvania  Engrg.  Co.,  1119-21  N.  Howard 

St,    Philadelphia,    and    315    Carpenter    Lane, 

Germantown,  Philadelphia,  Pa, 


OAKEY,  William  E.  (M  1932),  Consulting  Kngr., 

Oriakany,  N.  Y. 
OAKS,  Orion  0.   (M  1917),  Executive  B:ngr.t 

American  Radiator  Co.,  40  West  40th  St.,  New 

York,  N,  Y.,  and  (for  mail),  119  Oak  Ridge  Ave., 

Summit,  N.  J- 
GATES,  Walter  A,  (M  1931),  Htg.  and  Industrial 

Engr.,  Lynn  Gas  &  Electric  Co.,  90  Exchange 

St.,  and  (far  mail),  28&  Lynn  Shore  Dr.,  Lynn, 

Mass. 
O'BANNON,  Lester  $.*  (M  1028),  University  of 

Kentucky,  Lexington,  Ky, 
OBERGv  Harry  C.  (A  JTO),  Mgr,,  Kngrg,  Dept.f 

Crane  Co,,  Fifth  and  Broadway,  and  (for  mail), 

1362  W.  Minnehaha  St.,  St,  Paul,  Minn. 
OBERT,  Caflln  W.*  (M  1C16),  Consulting  Engr., 

Union  Carbide  &  Carbon  Research  Laboratories,  • 

IRC,,  30  East  42ncl  St.,  New  York,  and  (for  mail), 

m  N.  Columbus  Ave.,  Mt.  Vernon,  N.  Y. 
O'BRIEN,  J.  H.  (M  1923),  228  N.  LaSalk  St., 

Chicago,  111. 
Q'CONNELJU    Fresly   M.    (M    1016).   Resident 

En.gr.»  Inspector,  P.  W,  A.,  (for  mail),  College 

Court  Apts^  Pullman,  Wash. 
QFFUN,  Ben   (M  1028),  Owner  (for  mail).  B, 

Often  &  Co.»  308  S.  Dearborn  St.,  and  1100  N. 

Dearborn  St.,  Chicago*  111, 
OFFNER,  Alfred  J.*  (M  1922),  Consulting  Engr. 

(for  mall),  189  East  53rd  St,  New  York,  and 

I304Wltfa  Ave,,  Btechhurat,  U  1.,  N,  Y. 


O'GORMAN,  John  S.,  Jr.  (A  1934),  Sales  Engr. 

(for  mail),  Johnson  Service  Co.,  2142  East  10th 

St.,  Cleveland,  and  19205  Winalow  Rd.,  Shaker 

Heights,  Ohio. 
O'HARE,  George  W.,  Jr.  J  1935;  5  1032),  201 

West  72nd  St.,  New  York,  N.  Y. 
ORE,  William  C.  (J  1934),  Air  Cond.  Engr.  (for 

mail),  Sheldons,  Ltd.,  96  Grand  Ave.,  Gait,  and 

1200  Richmond  St.,  London,  Ont.,  Canada. 
OLCHOFF,   Maurice   (M   1933),   Mgr.,   Qlchoff 

Engrg.  Co.,  423  Dwight  Eldg.,  and  (for  mail), 

5341  Holmes,  Kansas  City,  Mo. 
OLSEN,  Carl  ton  F.  (A  1925;  J  1920),  Combustion 

Engr.,  Kewanee  Boiler  Co.,  Inc.,  1858  S.  Western 

Ave.,  and  (for  mail),  7914  Wabash  Ave.,  Chicago, 

111. 
OLSEN,  Gustav  E.  (M  1930),  GS09  Amstel  Blvd., 

Arverne,  L,  I.,  N.  Y. 
OLSON,   Bernhard   (A    1929),   122  S.  Michigan 

Ave.,  Chicago,  and  (for  mail),  5724  N.  Natoma 

Ave.,  Norwood  Park,  111. 
OLSON,  Gilbert  E.  (A/  1930),  440  Ward  Plcwy., 

Kansas  City,  Mo. 
OLSON,  Robert  G.   (M  1923),  Sales  Mgr,   (for 

mail),    Hydraulic   Coupling   Corp.,    Harper   at 

Russell,  and  HI  Putnam  Ave.,  Detroit,  Mich. 
OLVANY,  William  J.  (M  1912),  Prea.  (for  mail), 

William  J.  Olvany,  Inc.,  100  Charles  St.,  New 

York,  and  109-40~71st  Rd,r  Forest  Hills.,  L,  I,, 

N.Y. 
O'NEIL,  Joseph  M.  (A  1034),  332  Commonwealth 

Ave.,  Springfield,  Mass. 
O'NEILL,  James  W.  (M  1029;  A  1927;  J  1025), 

Chief  Engr.,  Trane  Co,  of  Canada,  Ltd.,  439 

King  St,  W.,  and  (for  mail),  8  Springmount  Ave,, 

Toronto,  Canada, 
O'NEILL,   Peter   (M  1920),   Treas,    (for   mail), 

Bartley-O'Neill    Co.,    240-42    Blvd.    of    Allies, 

Pittsburgh,  and  2448  Charles  St.   N.S.,   Pitta- 

burgh  (14),  Pa. 
OPPERMAN,  Everett  F,  (J  IQBfi;  ti  1TO),  169 

Milbank  Ave,,  Greenwich,  Conn. 
OREAR,  Andrew  G.  (M  1930),  Sales  Engr.  and 

Mfra.  Repr.  (for  mail),  Room  501,  San  Fernando 

Bldg,,  Los  Angeles,  and   1015  K.   Raleigh  St., 

Olendale,  Calif. 
O'RBAR,    L.   H.    (M   1934),    Pres,    (for   mail), 

Midwest  Pibg,  &  Htg,  Co.,  2450  Blake  St,,  and 

3033  West  37th  Ave.,  Denver,  Colo. 
OSBORN,    Wallace    J.    (A    liW7),    Vice-Prea,, 

Keency  Publishing  Co,,  Grand  Central  T«rm. 

Bid*.,  New  York,  N.  Y.r  and  (for  mail),  500  Old 

Post  Rd,,  Fairfield,  Conn. 
OSBORNE,  Gurdon  H.  (M  1922),  Gen,  Mgr,, 

The  Vtg.  &  Blow  Pipe  Co,,  Ltd,,  714  St,  Maurice 

St.,  Montreal,  and  (for  mail),  836  Pratt  Ave,, 

Outremont,  Montreal,  Que.,  Canada, 
QSBORNfc,  Maurice  M,  (M  1925),  387  Beacon 

St.,  Boston,  Mass. 
OSBURN,  Richard  M,  (J  1035:  S  1033),  2241 

Sedgwicfc  Ave,,  New  York,  N.  Y, 
OSTERLB,  William  H*   (M  1034),  Engr.   (for 

mall),  The  West  Penn  Electric  Co,,  W  Woocl  St., 

Pittsburgh,  and  333ft  Beacon  Hill  Ave.,  Dormant, 

Pa. 
OSTRIN*    Albert    (S   1835)»    1210   Jam«   N,, 

Minneapolis,  Minn, 
OTIS,  Gerald  E.*  (M  1022),  Vie^-Pm,  (for  mall), 

The  Herman  Nelson  Corp.,  and  1921-aSrd  Ave,, 

Mollne,  111. 
OTT,  Onm  W.  at  1025),  (Council  1984),  Con* 

suiting  Mech.  Eftgr.  (for  mail),  4^2  Wasbf&iton 

Bldg»»  and  123  S.  Virgil  Ave,f  toi  &jtgetts»  Calif, 
OTT,  Rush  C.  (M  ItBJ),  Sdei  Bagf.,  R« 

ating  Edulp,  Corp.,  9327  N*  Meridlun  St., 

napolis,  Ind. 
OUEUSOFF*  L.  S.  (M  1931), 

(for  mail),  Washtegton  Ga« 

St.  N.W..  Washington, 

St.,  Chevy  Chase,  Md* 
OVWTON,  Siduay  H*  (Jtf  IW0),  RMT,,  K,  .  V. 

Radiatore», 

land* 


Co. 


ROLL  OF  MEMBERSHIP 


PABST,  Charles  S.  (M  1934),  Pres.  and  Mgr. 
(for  mail),  Adams  Engrg.  Co.,  Inc.,  55  West  42nd 
St.,  New  York,  and  S727-98th  St.,  Woodhaven 
L.  I.,  N.  Y. 

PAETZ,  Herbert  E.  (M  1922),  Div.  Sales  Mgr. 
(for  mail),  American  Blower  Corp.,  2539  Wood- 
ward Ave.,  and  The  Ward  ell,  Detroit,  Mich. 

PAGE,  Harry  W.  (M  1923),  Pres.,  Wisconsin 
Equipment  Co.,  204  W.  Wisconsin  Ave.,  Mil- 
waukee, and  (for  mail),  7927  Warren  Ave., 
Wauwatoaa,  Wis, 

PAPPENFUS,  Wilfrid  G.  (S  1935),  312-13th 
Ave.  S.,  St.  Cloud,  and  (for  mail),  Pioneer  Hall, 
Minneapolis,  Minn. 

PARK,  Clifton  E).  (M  1929),  22  Otis  St.,  Need- 
ham,  Mass. 

PARK,  J.  Frank  (J  1930) f  Salesman  (for  mail), 
Carrier  Engrg.  Corp.,  748  E.  Washington  Blvd., 
Los  Angeles,  and  Route  3,  Box  956,  Modesto, 
Calif. 

PARKER,  Philip  (M  1915),  8  Middle  St.,  Woburn, 
Mass. 

PARROTT,  Lyle  George  (M  1922),  Consulting 
Engr.,  McColl,  Snyder  &  McLean,  2308  Pen- 
obscot  BldK.,  and  (for  mail),  4078  Seebaldt  Ave., 
Detroit,  Mich. 

PARSONS,  Roger  A.  (J  1933),  Sales  Engr.,  Dail 
Steel  Products  Co.,  and  (for  mail),  525  W.  Grand 
River  Ave.,  Lansing,  Mich. 

PARTLAN,  James  W.  (Life  Member,  M  1916), 
14290  Goddard  Ave,,  Detroit,  Mich. 

PATERSON,  James  S.*  (M  1922),  Mech.  Engr. 
(for  mail),  Board  of  Education,  155  College  St., 
and  23  Norton  Ave.,  Toronto,  Ont.,  Canada. 

PATORNO,  Sullivan  A.  S.  (U  1923),  Chief 
Draftsman  (for  mail),  Meyer,  Strong  &  Jones, 
Inc.,  101  Park  Ave,,  and  312  East  163rd  St., 
New  York,  N.  Y, 

PAUL,  Donald  I.  (J  1932),  Sales  Engr.  (for  mail), 
Gurney  Foundry  Co.,  Ltd.,  4  Junction  Rd.,  and 
222  Kern  Ave.,  Toronto,  Ont.,  Canada. 

PEACOCK,  James  K.  (M  1921),  Hoffman 
Specialty  Co.,  Inc.,  500  Fifth  Ave.,  New  York, 
and  (for  mail),  440  Fowler  Ave.,  Pelliam  Manor, 

PEEBLES,  John  K.,  Jr.  (A   1925;  J  1924),  7 

Brandon  Apts.,  University,  Va. 
PBLLBR,  Leonard  (J  1934),  Engr,,  D.  J.  Peller 


and  4209  Grove  Ave.,  Richmond,  Va, 

PENNEL,  Reed  (J  1033)  ,  1335  Grand  Ave.,  St. 
Paul.  Minn. 

PENNOCK,  William  B,  (M  1927),  Dial.  Sales 
Engr,,  T»n«  Co,  of  Canada,  Ltd.,  and  (for  mail), 
02  Markland  St.,  Hamilton,  Ont,,  Canada, 

PBRINA,  Arthur  E,  (/  1035;  5  1933),  126  Court- 
land  St.,  Staten  Island,  N.  Y.,  and  (for  mail), 
Box  198,  Carnegie  Institute  of  Technology, 
Pittsburgh,  Pa. 

PERKINS,  Robert  G.  (A  1935),  Sales  Engr.,  llg 
Electric  Ventilating  Co.,  Chicago,  111,,  and  (for 
mail),  1&87  N*  Parkway,  Memphis,  Tenn, 

PBSTERPIELD,  Charles  H,  (S  1982),  Box  554, 
University  Station,  Dept  of  Mech.  Enarg., 
University  of  North  Dakota,  Grand  Forks,  N,  D. 

PETERS,  Herbert  H.  (M  1080),  Mar.,  H.  H. 
Peteti  Heating  Co.,  1842  North  40th  St.,  Mil* 


. 

PETERSON*  SterUnflD.  (A  1930),  Br.  Mgr,  (for 
mall),  Johnson  Service  Co.,  473  Colman  Blag,, 
and  50Bl  Prince  St.,  Seattle,  Wash. 
PFB1FER,  Otto,  Jr.  (A  1035;  J  1932),  Bagr,. 
Ralph  D.  Thonaai  &  Associates,  1200  Second 
S.,  ftnd  (for  mail),  1516  Monroe  St*  N,E,, 
olis,  Minn, 

,   John  F*   (M  1980»  J  1925),  $46 
Louisa  St*,  WlUtawtport,  Pa. 
PFUHL&R,  Johxi  t.  (A  1926]  J  1023),  Plbg.  and 
Ht£,  000  Manor  Rd.,  West  New  Brighton,  S.  I., 

PHILIP,  WtiMfcm  (M  WO),  74  Battedo  Am, 
Toronto,  Ont,,  Canada, 


PHILLIPS,  Frederic  W.,  Jr.  (U  1921),  L.  J. 
Mueller  Furnace  Co.,  101  Park  Ave.,  New  York, 
and  (for  mail),  825  East  SSth  St.,  Brooklyn,  N.  Y. 

PHIPPS,  Frederick  G.  (M  1930),  Vice- Pres., 
Preston  Phipps,  Inc.,  955  St.  James  St.  W.,  and 
(for  mail),  2054  Mercier  Ave.,  Montreal,  P.  Q., 

PIERCE,'  Edgar  D.  (J  1933),  Engr.,  Carrier 
Engrg.  Corp.,  748  E.  Washington  Blvd.,  and  (for 
mail),  360  West  68th  St.,  Los  Angeles,  Calif. 

PIERCE,  William  MacL.  (J  1935;  S  1933), 
Research  (for  mail),  Mine  Safety  Appliance  Co,, 
Braddock,  Thomas  and  Meade  Sts.,  Pittsburgh, 
Pa.,  and  31  Potter  St.,  Melrose,  Mass. 

PIHLMAN,  Arthur  A.  (M  1928),  (for  mail), 
Consolidated  Gas  Co.  of  New  York,  4  Irving  PL, 
New  York,  N.  Y.,  and  235  Dwight  St.,  Jersey 
City,  N.  J. 

PILLEN,  Harry  A.  (A  1933),  Mfg.  Repr.  (for 
mail),  Harry  A,  Pillen  Co.,  622  Broadway,  and 
2208  Crane  Ave.,  Cincinnati,  Ohio. 

PINDER,  Percy  H.  (M  1919),  366  Third  Ave., 
New  York,  N.  Y. 

PINES,  Sidney  (M  1920),  Vice-Pres.  (for  mail), 
Natkin  &  Co.,  2020  Wyandotte  St.,  and  5225 
Charlotte  St.,  Kansas  City,  Mo. 

PISON,  Donate,  Jr.  (J  1935;  S  1933),  Moloiloilo, 
Phillipine  Islands. 

PISTLER,  William  C.  (M  1934),  Mech.  Engr.  in 
Charge  of  Design,  Carl  J.  Kiefer,  Consulting 
Engr.,  91S  Schmidt  Bldg,,  and  (for  mail), 
Orchard  Lane  and  Crestview  Ave.,  Pleasant 
Ridge,  Cincinnati,  Ohio. 

PITCHER,  Lester  J.  (M  1929;  A  1928;  J  1924), 
8129  Dante  Ave.,  Chicago,  111. 

PITTOCK,  Louis  B.  (M  1930),  (for  mail),  429- B 
Oliver  Bldg.,  and  80  Berry  St.,  Crafton  Station, 
Pittsburgh,  Pa. 

PIZIE,  Stuart  G.  (A  1926),  215-17  N.  Flagler  Dr., 
West  Palm  Beach,  Fla. 

PLACE,  Clyde  R.  (M  1924) ,  Consulting  Engr.  (for 
mail),  420  Lexington  Ave.,  and  333  East  57th  St., 
New  York,  N.  Y. 

PLAENERT,  Alfred  B.  (A  1933;  J  1927),  1102 
S.  Park  St.,  Madison,  Wis, 

PLASS,  Charles  Webster  (M  1928) ,  826  E,  Haines 
St.,  Philadelphia,  Pa. 

PLAYFAIR,  George  Alexander  (A  1924),  Mgr, 
(for  mail),  Johnson  Temperature  Regulating  Co. 
of  Canada,  Ltd.,  97  Jarvis  St.,  Toronto,  and  West 
Hill,  Ont.,  Canada. 

PLEWES,  Stanley  E.  (M  1917),  Philadelphia 
Mgr,  (for  mail),  Johnson  Service  Co.,  2853  North 
12th  St.,  North  Philadelphia  Station  8,  Phila- 
delphia, and  309  Evergreen  Rd.,  Jenkmtown,  Pa. 

PLUM,  Leroy  H.  (M  1934),  Industrial  Engr., 
Minneapolia-Honerwell  Regulator  Co.,  2240  N. 
Broad  St.,  Philadelphia,  Pa.,  and  (for  mail), 
216  Guilford  Ave,,  ColHngswood,  N.  J. 

PLtJNKETT,  John  H.  (M  1925),  81  Woodrow 
Ave.,  Boston,  Mass. 

POEHNBR,  Robert  E.  (M  1928),  Vice-Pres-Secy,. 
W.  H.  Johnson  &  Son  Co.,  330  E.  St,  Joe  St.,  and 
(for  mail),  2308  Coyner  Ave.,  Indianapolis,  Ind, 

POHLB,  K,  F.  (A  1930),  Vice-Pres.,  W.  F.  Hirach- 
man  Co.,  Inc.,  202  East  44th  St.,  New  York, 
N.  Y, 

POLDERMAN,  Lambert  H.  (M  1927),  Vice- 
Pres.  (for  mail),  Carrier  Engrg.  Corp.  of  Cali- 
fornia, 748  E,  Washington  Blvd.,  and  3462 
Lambeth  St.,  Los  Angeles,  Calif. 

POLLARD,  Alfred  L*  (A  1032),  Gen.  Supt., 
Steam  Heat  Dept.  (for  mail),  Puget  Sound 
Power  &  Light  Co.,  601  Electric  Bldg*,  and  3009 
28th  W.,  Seattle,  Wash. 

POPE,  S.  Austin  (M  1917),  Prea.  (for  mail), 
William  A,  Pope  Co.,  26  N,  Jefferson  St..  Chicago, 
and  831  Ashland  Ave.,  River  Fore&tf  III, 

PORTER,  Herbert  M.  (U  1981),  65  North  17th 
St.*  Minneapolis,  Minn. 

POSEYt  James  (M  1,919),  Consulting  Enor.  (for 
mail),  175S  Baltimore  Trust  Bldg.,  and  4005 
Uberty  Heights  Ave.,  Baltimore,  Md. 

POTVlNt  L«o  J.  (A  1934).  Sales  Bngr.,  Hoffman 
Specialty  Co,,  Inc.,  ISO  N.  Wells  St,,  Chicago, 
and  (for  mail),  341  Walnut  St.,  Eimhurst,  111, 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


POUCHER,  Richard  C.  (S  1935),  1480  Chelms- 

ford  St.,  St.  Paul,  Minn. 
POWELL,  Knox  A.  (J  1935;  5  1933),  1008-lSth 

Ave.  S.E.,  Minneapolis,  Minn. 
POWERS,  Edgar  G.  (A  1934;  /  1931),  (for  mail), 

James  A.  Walsh,  Inc.,  Architects  Bldg.,  17th  and 

Sansom  Sts.,  Philadelphia,  Pa.,  and  304  Crest 

Ave.,  Haddon  Heights,  N.  J. 
POWERS,  Fred  I.  (M  1920),  Factory  Repr.  (for 

mail),  Box  324,  and  605  S.  Sixth  Ave.,  Bozeman, 

Mont. 
POWERS,  Fred  W.   (M  1911),  Pres.  and  Gen. 

Mgr.    (for   mail),   The   Powers   Regulator   Co., 

2720  Greenview  Ave.,  and  900  Castlewood  Ter- 
race, Chicago,  III. 
POWERS,  Lowell  G.  (J  1930),  Sales  Engr.  (for 

mail},  Carrier  Engrg.  Corp.,  1501  Carew  Tower, 

Cincinnati,  Ohio,  and  325  W.  Diamond  Ave., 

Hazleton,  Pa. 
PRENDERGAST,    James    J.    (S    1934),    2114 

Stearns  Rd.,  Cleveland,  Ohio. 
PRENTICE,  Oliver  J.  (A  1927),  (for  mail),  C.  A. 

Dunham  Co.,  450  E.  Ohio  St.,  and  850  Lake 

Shore  Dr.,  Chicago,  111. 
PRESDEE,  CHff  W.  (A  1920),  Adv.  Mgr.,  Heating 

&  Ventilating,   148  Lafayette  St.,  New  York, 

N.  Y. 
PRICE,  Charles  E.  (A  1933),  Treas.  (for  mail), 

Keeney  Publishing  Co.,  G  N.  Michigan  Ave., 

Chicago,  and  1151  ChatEeld  Rd.,  Winnetka,  111. 
PRICE,  D.  O.  (M  1934),  Htg.  and  Air  Cond.  Engr., 

General  Steel  Wares,  Ltd.,  199  River  St.,  and 

(for  mail),  145  Eastbourne  Ave.,  Toronto,  Out,, 

Canada. 
PRICE,  Ernest  H.   (J  1934;  5  1932),  c/o  Haff 

Supply,  Inc.,  Box  328,  Riverhead,  L.  I.,  N.  Y. 
PRIESTER,  Gayle  B.  (S  1934),  814  Fulton  St. 

S.E.,  Minneapolis,  Minn. 
PRYIBIL,  Parul  L.   (A  1932),  Partner,  Hucker- 

Pryibil  Co.,  1700  Walnut  St.  Philadelphia,  and 

(for  mail),  328  E.  Philellena  St.,  Germantown, 

Philadelphia,  Pa. 
PRYOR,    Frederick  L.   (A/   1913),   5   Colt  St., 

Paterson,  N.  J. 
PUNG,  Donald  W.  (S  1935),  709  Ninth  Ave.  N., 

St.  Cloud,  and  (for  mail),  315-lGth  Ave.  S.E., 

Minneapolis,  Minn. 
PURCELL,  Frederick  C.  (M  1020),  Dial.  Mgr. 

(for  mail),  National  Regulator  Co.,  2847  Grand 

River  Ave.,  and  18680  Santc  Rosa  Dr»,  Detroit, 

Mich. 
PURCELL,  Robert  E.  (M  1916),  4001  Seebaldt 

Ave.,  Detroit,  Mich. 
PURDY,  A.  K.  (M  1922),  Pres.  (for  mail),  Purdy 

Manaell,  Ltd.,  63  Albert  St.,  and  30  Glenrose 

Ave.,  Toronto,  Ont.»  Canada. 
PURDY,  Randall  B.   (A   1927),  Assoc,  Editor, 

Power  (for  mail),  McGraw-Hill  Publishing  Co., 

330  West  42nd  St.,  New  York,  and  224~Q5-139th 

Ave.,  Laurelton,  L.  I.,  N.  Y. 
PURINTON,  Dexter  J.  (A  1923),  Associate  (for 

mail),  Voorhees,  Gmelin  &  Walker,  Archts,,  101 

Park  Ave,,  New  York,  N,  Y.,  and  23  Sachem 

Rd.,  Greenwich,  Conn. 
PURSELL,    H.    E,    (A/    1919),    Special    Repr., 

Kewanee  Boiler  Corp,,  Kewanee,  IU. 
PYLE,  John  W,  (M  1919),  Peru  Htg,  Co.,  30  W. 

Canal  St.,  Peru,  hid, 


QUAY.  0.  M.*  (Charter  Member;  Life  Member,* 
Presidential  Member),  (Pres.,  1900;  1st  Vke- 
Pres,,  18064899;  2nd  Vice-Pres.,  1895),  725 
Eastern  Ave.,  Belief ontaine,  Ohio, 

QUEER,  Elmer  Roy  (M  1933),  Research  Engr. 

gor   mail),    The    Pennsylvania   State   College 
ngrg,  Experiment  Station,  and  Arbor  Way, 
State  College,  Pa. 

QUINUVAN,  Lawrence  P»  (J  1935;  3  1933), 
As8t.»  Case  School  of  Applied  Science,  and  (for 
mall),  1S7U  Earl  wood  Rd.,  Cleveland,  Ohio. 
QUIGLEY,  William  J.  (M  1920),  27  JCnowKoa 
Ave*,  Kenmore,  N,  Y. 


QUIRK,  Clinton  H.  (M  191<>;  J  1015),  Sales 
Engr.  (for  mail1),  Vento  Div.,  American  Radiator 
Co.,  40  West  40th  St.,  New  York,  and  405  Front 
St.,  Hempstead,  L.  L,  N.  Y. 

R 

RACHAL,  John  M.  (J  1930),  Mgr.,  Air  Cond. 
Dept.  (for  mail),  Carrier-Brunswick  Inter- 
national, Inc.,  850  Frelinghuysen  Ave.,  Newark, 
and  01  S.  Mimn  Ave.,  East  Orange,  N.  J. 

RACK,  Edgar  C.  (U  1931),  Consulting  Engr., 
Johns-Manville,  22  East  40th  St.,  New  York, 
N.  Y.,  and  (for  mail),  2SS  Park  Ave.,  East 
Orange,  N.  J. 

RAFFES,  Abraham  (J  1935;  5  1932),  care  of 
I.  Pontak,  977  East  178th  St.,  New  York,  N.  Y. 

RAINE,  John  J.  (M  1912),  Vice-Pros,  (for  mail), 
The  G.  S.  Blodgett  Co.,  190  Bank  St.,  and  Essex 
Jet.,  V.  P.,  Burlington,  Vt. 

RAINGER,  Wallace  F.  (A  1930;  J  1924),  441 
Hawthorne  Ave.,  Yonkers,  N.  Y. 

RAISLER,  Robert  K,  (A  1933;  J  1930),  Treaw. 
(for  mail) ,  Raisler  Htg.  Co.,  329  Amsterdam  Ave., 
and  25  East  77th  St.,  New  York,  N.  Y. 

RAMSEY,  Raymond  F,  (5  1933),  1522  Coutaht, 
Lakewood,  Ohio. 

RANCKT  Guy  L.  (A  1933),  MRF.,  C.  A.  Dunham 
Co.,  3605  Laclede  Ave.,  St.  Louis,  and  (for  mail), 
472  Pasadena  Ave.,  Webster  Groves,  Mo, 

RANDALL,  W.  Clifton*  (M  1928),  Detroit  Steel 
Products  Co.,  2250  E.  Grand  Blvd.,  Detroit, 
Mich. 

RANDOLPH,  Charles  H.  (A*  1930;  ,4  1028; 
J  1926),  Air  Cond.  En«r.,  The  Milwaukee 
Electric  Railway  &  Light  Co.,  217  W.  Michigan 
St.,  and  (for  mail),  1925  N.  Prospect  Ave., 
Milwaukee,  Wis. 

RASMUSSEN,  Robert  P.  (fl/  1931),  Prea. 
Economy  Equipment  Co.,  0835  Weutwwth  Ave., 
and  (for  mail),  1243  East  40th  St.,  Chicago,  lit. 

RATHBUN,  Perry  W.  (M  1033),  Resident  Kngr. 
Inspector  P.  W,  A.,  and  (for  mail),  1H09  North- 
west 37th  St.,  Oklahoma  City,  Okliu 

RATHER,  Max  F.  (M  1919),  Johnson  Service  Co., 
2142  East  19th  St.,  Cleveland,  Ohio. 

RAUH,  Edward  M.  (S  1934),  (for  mail),  205  K. 
Boyd,  Norman  and  Alva,  OkUi, 

RAY,  Lewis  B.  (M  1932),  Frew,  (fur  mail),  Ray 
Engrg.  Co.,  Inc.,  800  Broad  St.,  Newark,  and 
151  Augusta  St.,  Irvington,  N.  J. 

RAYMER,  William  F.,  Jr.  (J  1934),  Sales  Kngr. 
(for  mail),  American  Blower  Corp.,  402  Brtmd 
st,,  Newark,  and  50  N.  Mtmn  Ave.,  Eait  Orange, 
N.J. 

RAYMOND,  Fred  L*  (A  1020),  PreE.  (for  mail), 
F*  L  Raymond  Co,,  620  W.  Washington  Blvd., 
Chicago,  and  547  N.  Keystone  Ave,,  River 
Forest,  111. 

RAYN1S,  Theodore  (J  1934),  Ami,  Supervisor, 
Htg.  and  Vta.,  New  York  Navy  Yard  Central 
Drafting  Ofhcc,  Vent.  Sect.,  and  (for  mail), 
8631-79th  St.,  woodhavcn,  L.  L,  N.  Y. 

READ,  Robert  R.  (S  1034),  (for  mall),  1975 
Taylor  Rd,,  East  Cleveland,  and  27aa  Owufea 
Rd,,  Cuyahoga  Falls,  Ohio. 

RECK,  WUliftm  E.  (M  1927),  Civil  Engr,,  The 
Reck  Heating  Co.,  Ltd.,  Esromgade  1$,  Copen- 
hagen N,,  ana  Sundvey  10,  Hdlerup,  Denmark* 

RBDF1&LD,  Clarke  (J  1935;  5  1982),  318  Hngle 
St.,  Tenafly,  N.  J, 

REDSTONE,  Arthur  L.  (M  1931),  Research 
Engr.  (for  mail),  Proctor  8t  Schwartz  Seventh 
and  Tabor  Rd..  and  Park  Towers,  Kmbte  and 
Ogonta  Ave.,  Philadelphia,  Fa. 

REED,  Irvij*&  <3,  (/  1&34),  A«at.  Supt  and  Chief 
Engr,  (for  mail),  Grant  BldK»,  Inc.,  Room  417, 
310  Grant  St.»  and  8009  Homt  Ave,,  Mt.  Oliver, 
Pittsburgh,  Pa. 

REED,  John  F,  (M  1027:  A  1023),  Vioa*Ptm  (for 
mail),  American  Air  Filter  Co,,  420  ," 
Ave,,  New  York,  and  87  Sapmore  T 
vWe,  N,  Y, 


32 


ROLL  OF  MEMBERSHIP 


REED,  Paul  L.  (A  1932),  1034  Art  Hill  PI.,  St. 

Louis,  Mo. 
REED,  Van  A.,  Jr.  (M  1930),  Mech.  Engr.  (for 

mail),  Federal  Engineering  Co.,  239  Fourth  Ave., 

Pittsburgh,  and  114  Water  St.,  Elizabeth,  Pa. 
REED,    William    M.    (M    1927),    American   Air 

Filter  Co.,  215  Central  Ave.,  Louisville,  Ky. 
REGER,  Henry  P.  (M  1934),  Pres.  and  Treas.  (for 

mail),  H.  P.  Reger  &  Co.,  1501  East  72nd  PL, 

and  6939  Bennett  Ave.,  Chicago,  111. 
REID,  Henry  P.  (M  1981;  A  1927),  Special  Engr. 

(for  mail),  Universal  Atlas  Cement  Co.,  20S  S. 

LaSalle  St.,  Chicago,  and  3507  Oak  Park  Ave., 

Berwyn,  111. 
REID,  Herbert  F.  (A  1932),  Reid-Graff  Plbg.  Co., 

1417  Peck  St.,  Muskegon  Heights,  Mich. 
REIIXY,  Charles  E.   (J  1928),  4920  City  Line 

Ave.,  Philadelphia,  Pa. 
REIJLLY,  J.  Harry  (M  1931;  A  1931;  J  1929), 

Sales  Engr.,  American  Radiator  Co.,  402  Broad 

St.,  Newark,  and  (for  mail),  14  Watson  Ave., 

East  Orange,  N.  J. 
REINKE,  Alfred  G.  (J  1933),  Group  Leader  on 

Instruments,  Westinghouse  Electric  &  Mfs.  Co., 

95  Orange  St.,  Newark,  and  (for  mail),  319  Park 

PI.,  Irvington,  N.  J. 
RENOUF,  E.  Prince  (M  1933),  Mar.,  Air  Cond. 

Dept.   (for  mail),  Straus  Frank  Co.,  and  1901 

MacGrcgor,  Houston,  Texas. 
RENTE,  Harry  W.   (M  1931),  IltR.  Engr,,  Oil 

Burners,  70  W.  Clnppcwa  St.,  and  (ior  mail),  114 

Morris  Ave.,  Buffalo,  N.  Y. 
RENTE,  Sidney  R.  (A  1930),  31  Garrison  Rd., 

WiWamsville,  N.  Y. 
REI>KO,  Joseph  J,  (S  1034),  4024  Ilamm  Ave., 

Cleveland,  Ohio. 
RETTKW,  Harvey  F.  (A/  1929),  Htg.  anc^Vtfl. 

JKngr.,   Board  ot   Education,  21st  and  Winter, 

and  (for  mail),  6821  Martins  Mill  Rd,,  Phila- 
delphia, Pa, 
REYNOLDS,  Jack  A.   (J  1035;  S  1933)    Asst. 

Engr.,  Sherman  Mf#.  Co.,  ana  (for  mail),  810 

Fischer  Av«.,  Sherman,  Texas. 
REYNOLDS,  Thurlow  W.  (M  11)22),  Consulting 

Entjr.,  100  Pinecrest  Dr.,  Hawtings-on-IIudson, 

N,  Y. 
REYNOLDS,  Walter  V.  (A  1928),  Pres.,  Walter 

Reynolds,  Inc.,  8G1  Third  Ave,,  New  York,  N,  Y. 
RHEA,  Chester  A.  (A  1031),  Steel  Boiler  Renr., 

National  Radiator  Corp.,  21M4  Arch  St.,  and  (for 

mail),  722  Carpenter  Lane,  Philadelphia,  Pa. 
RICE,  C.  J.  (A  1023),  Pres.  (for  mail),  Sterling 

Kngrg,  Co.,  3738  N.  Holton  vSt.,  und  3370  N. 

Summit  Ave.,  Milwaukee,  Wis. 
RICE,  Robert  B.  (M  1934),  Aaaoc.  Prof,  in  Mech. 

Engrg.  (for  mail),  Newark  College  of  Engineering, 

307  High  St.,  Newark,  and  105  Rutgers  PL, 

Nutley,  N,  J, 
RICHARD,  Edwin  J,  (M  1033),  Owner  (for  mail). 

Edwin  J.  Richard  Equipment  Co.,  Chamber  of 

Commerce  Bldg,»  and  3504  Paxton  Ave.,  Cincin- 
nati, Ohio. 
RICHARDSON,  Heitry  G.  (M  1034) »  Vice-Press,, 

Hawfey,  Richardson,  Williams  Co.,  204  Cooly 

Bldg.,  and  (for  mail),  1433  Harvard  Ave.,  Salt 

Lake  City,  Utah. 
RICHARDSON*  Henry  Thomas  (A  1930),  Vice- 

Prea.  (for  mail),  Richardson  Sc  Boynton  Co,,  244 

Madison  Ave.,  and  156  East  79th  St.,  New  York, 

N.  Y. 
RICHMOND,  Jotm  (S  1933),  5035  Forbes  St., 

Pittsburgh,  Pa, 
RtCHTMANN,  WilHam  M.*  (A  1932;  J  1920). 

Asst.  Prof,  of  Bngrg,  (for  mail),  Texas  College  of 

Arts  and  Industries,  and  709  W,  Santa  Gertrudes 

St.,  Kingsvllle,  Texm 
RI00LE,  Kemfole  L,   (.7  1985;  S  1033),  4150 

Windsor  St,  Pittsburgh,  Pa, 
RIBS,  Lester  S.  (M  1929),  Asst.  Stipt.  of  Bldga, 

and  Grounds,  University  of  Chicago,  060  East 

5Sth  St.,  and  (for  roaU)»  $614  Bkckstone  Ave.» 

Chicago,  III, 


RIESMEYER,  Edward  H.,  Jr.  (J  19301,  Ht?. 
Engr.,  Schaffer  Htg.  Co.,  231-33  Water  St.,  and 
(for  mail),  4702  Stanton  Ave.,  Pittsburgh,  Pa. 

RIETZ,  Elmer  W.*  (M  1923),  Gen.  Sales  Mgr.  (for 
mail),  Powers  Regulator  Co.,  2720  Green  view 
Ave.,  Chicago,  and  940  Greenwood  Ave., 
Winnetka,  111. 

RILEY,  Champlain  L.  (M  1906),  (Presidential 
Member],  (Pres.,  1921;  1st  Vice-Pres.,  1920; 
Council,  1918-1922),  Clark,  MacMullen  &  Riley, 
Inc.,  101  Park  Ave.,  New  York,  N.  Y. 

RILEY,  Edward  C.  (J  1935;  S  1933),  Research 
Worker,  Harvard  School  of  Public  Health,  55 
Shattuck  St.,  Boston,  and  (for  mail),  51  Centre 
St.,  Brookline,  Mass. 

RILEY,  Robert  G.  (S  1934),  8S-37-179th  St., 
Jamaica,  N.  Y. 

RINEHARD,  Wilson  R.  (J  1932),  Choudrant, 
La. 

RITCHIE,  A.  Gordon  (M  1933),  Pres.  and  Mgr. 
(for  mail),  John  Ritchie,  Ltd.,  102  Adelaide  St. 
E.,  and  41  Garfield  Ave.,  Toronto,  Canada. 

RITCHIE,  Edmund  J.  (M  1923),  yice-Pres., 
(for  mail),  Sarco  Co.,  Inc.,  183  Madison  Ave., 
New  York,  and  2  Grace  Court,  Brooklyn,  N.  Y. 

RITCHIE,  William  (M  1909),  Vice-Pres.,  Boyn- 
ton Furnace  Co.,  373  Fourth  Ave.,  New  York, 
N.  Y.,  and  (for  mail),  17  Van  Reipen  Ave., 
Jersey  City,  N.  J. 

RITTER,  Arthur  (M  1911),  New  York  Dist.  Mgr. 
(for  mail),  American  Blower  Corp.,  401  Broad- 
way, New  York,  and  29  Edgemont  Rd.,  Scars- 
dale,  N.  Y. 

ROBB,  John  M.  (M  1913),  Consulting  Engr., 
1513  Columbia  Terrace,  Peoria,  III. 

ROBERTS,  Henry  L.  (M  1916),  Htg.  Engr.  and 
Contractor  (for  mail),  Henry  L.  Roberts,  228 
North  16th  St.,  Philadelphia,  and  1014  AHston 
Rd.,  Brookline,  Delaware  Co.,  Upper  Darby 
P.  O.,  Pa. 

ROBERTS,  James  R.  (J  1934),  Engr.  (for  mail), 
Sutherland  Air  Cond.  Corp.,  627  Marquette 
Ave.,  and  2423  Portland  Ave.  S.,  Minneapolis, 
Minn. 

ROBINSON,  Harry  C.  (M  1930),  Htg.  Engr.,  07U 
Plcaaant  St.,  Worcester,  Mass. 

ROCKWELL,  Theodore  F.  (M  1933;  A  1933; 
J1932),  Instructor  in  Htg.  and  Vtg.  (for  mail), 
Carnegie  Institute  of  Technology,  and  131 
Edgewood  Ave.,  Edgewood,  Pittsburgh,  Pa. 

RO0ENHEISKR,  Georfte  B.  (M  1933),  Head, 
Htg.  and  Vtg.  Dept.  (for  mail),  David  Ranken, 
Junior  School  of  Mech.  Trades,  4431  Finney  Ave., 
and  3639  A;  Dover  PL,  St.  Louis,  Mo. 

R.OIXSERS,  Frederick  A.  (A  1934),  Br.  Mgr., 
Minneapolis-Honeywell  Regulator  Co.,  4500 
Euclid  Ave.,  Cleveland,  and  (for  mail),  2577 
Ashton  R,d.»  Cleveland  Heights,  Ohio. 

RODGERS,  Joseph  S.  (/  1034),  Engr.,  Mont- 
gomery Ward  &  Co.,  1000  S.  Monroe  St.,  Balti- 
more, and  (for  mail),  1  Third  Ave.,  Brooklyn 
Park,  Md. 

RODMAN,  Robert  W,  (M  1922),  Supt.  of  Plant 
Operation  (for  mail),  Board  of  Education,  City 
of  New  York,  500  Park  Ave.,  and  175  West  73rd 
St.,  New  York,  N,  Y.  * 

ROEBUCK,  William,  Jr.  (M  1D17),  Mfrs.  Repr. 
(for  mail),  311  Jackson  Bldg.,  and  154  Sanders 
Rd,,  Buffalo,  N.  Y, 

ROHX/IN,  Karl  W.  (M  1930),  Engr.,  Warren 
Webster  &  Co.,  17th  and  Federal  Sts,,  Caraden, 
and  (for  mail),  4453  Terrace  Ave.,  Merchant- 
vffle,N.  J* 


RO3ULANI>,  S.  L.  (A  1934),  Design  Engr.,  Okla- 
homa Gas  &  Electric  Co.,  Oklahoma  City,  Okla. 

ROSE,  Howard  J.  (M  1984),  Sales  Engr,,  Fitz- 
gibbons  Boiler  Co.,  Inc.,  18&  Main  St.,  White 
Plains,  and  (for  mail),  100  Siebrecht  PL,  New 
Rochelta,  N.  Y, 

ROSEBERRY,  John  H.  (U  1931),  32  Wardman 
Rd.,  Kenmore,  N,  Y. 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


ROSEBROUGH,  Robert  M.  (M  1920),  Br.  Mgr. 
(for  mail),  L.  J.  Mueller  Furnace  Co.,  4246 
Forest  Park  Blvd.,  and  6012  McPherson  Avc,, 
St.  Louis,  Mo. 

ROSELL,  Axel  F.  (M  1035),  Mgr.  Gen  Sales 
Dept.,  Svenska  Flaktfabriken  Kungsgatan  S, 
Stockholm,  and  (for  mail),  Ku  Atlas  0,  Lidingo, 
Sweden. 

ROSENBERG,  Philip  (A  1928),  Secy-Treas., 
Universal  Fixture  Corp.,  137  West  23rd  St.,  and 
(for  mail),  250  West  104th  St.,  New  York,  N.  V. 

ROSS,  John  O.*  (M  1920),  Ross  Industries  Corp., 
350  Madison  Ave.,  New  York,  N.  Y. 

ROSSITER,  Paul  A.  (S  1935),  SS  Kent  St., 
Minneapolis,  Minn. 

ROTH,  Charles  F.  (A  1930),  Mgr.,  International 
Htg.  and  Vtg.  Exposition,  Grand  Central 
Palace,  New  York,  and  (for  mail),  141  East  36th 
St.,  New  York  (November  1  to  April  30),  and 
Dreamthorp,  Bedford  Village,  N.  Y.  (May  1  to 
October  31). 

ROTH,  Harold  Raymond  (M  1935),  Mgr. 
Toronto  Office  (for  mail),  Canadian  Sirocco  Co., 
Ltd.,  Room  321,  57  Bloor  St.  W.,  and  18 
Tichester  Rd.,  Toronto,  Ont,  Canada. 

ROTTMAYER,  Samuel  I.  (.4  1933;  J  1928), 
Mech.  Engr.  (for  mail),  Samuel  R.  Lewis,  407 
S.  Dearborn  St.,  and  1109  Hyde  Park,  Blvd., 
Chicago,  111. 

ROWE,  William  A.  (M  1921),  (Council,  1920- 
1931),  718  Longfellow  Ave.,  Detroit,  Mich. 

ROWLEY,  Frank  B.*  (U  1918),  (Presidential 
Member},  (Pres.,  1932;  1st  Vice-Pres.,  1931;  2nd 
Vice-Pres.,  1930;  Council,  1927-1933),  Prof,  of 
Mech.  Engrg.  and  Director  of  Experimental 
Engrg.  Lab.,  University  of  Minnesota,  and  (for 
mail),  4801  E.  Lake  Harriet  Blvd.,  Minneapolis, 
Minn. 

ROYER,  Earl  B.  (M  1028),  Designing  Engr., 
Fosdick  &  Hilmer  Consulting  Engrs.,  1703  Union 
Trust  Bldg.,  and  (for  mail),  603.5  Iris  Ave., 
Cincinnati,  Ohio. 

ROZETT,  William,  Jr,  (J  1935;  S  1932),  3528  E. 
Tremont  Ave.,  New  York,  N.  Y. 

RUDIO,  H.  M.  (M  1921),  Mgr.,  Air  Cond.  Dept. 
(for  mail),  Gustin  Bacon  Mfg.  Co.,  1412  West 
12th  St.,  and  0039  Kdgerale  Rd.,  Kansas  City, 
Mo. 

RUFF,  DeWltt  O.  (M  1922),  Heal y- Ruff  Co,,  765 
Hampden  Ave.,  St.  Paul.  Minn. 

RUGART,  Karl  (A  1924),  Dist.  Mgr.  (for  mail), 
Warren  Webster  &  Co.,  20  South  20th  St.,  and 
5830  Willows  Ave.,  Philadelphia,  Pa. 

RUPPERT,   Edward   H.    (A    1923),  85  Eastern 

1    Pkwy.,  Brooklyn,  N,  Y. 

RUSSELL,  Joseph  Nelson  (M  1890),  Managing 
Dir.  (for  mail),  Rosaer  &  Russell,  Ltd,,  Romney 
House,  Marsham  St.,  Westminster,  and  Per- 
nacres  Fulmer  near  Slough,  Buckinghamshire, 
England, 

RUSSEIX,  W.  A.  (M  1921),  (Council,  1934),  M«r.» 
K.  C.  Br,  (for  mail),  U.  S,  Radiator  Corp.,  1405 
West  llth  St.,  and  230  Ward  Pkwy.,  Kansas 
City,  Mo, 

RUSSEIX,  William  B.  (M  1928).  Colorado  Ave., 
R,  F,  D.  No,  1.,  Jolict,  111. 

RYAN,  Marry  J,  (M  1022),  47  Harris  Ave., 
Albany,  N.  Y. 

RYAN,  William  F.  '</  1983),  Sales  Engr.,  Lee 
Hardware  Co,,  252*54  N.  Santa  Fe,  and  (for 
mail),  600  E.  Iron,  Salina,  Kans. 

RY0ELL,  Carl  A.  (M  1931;  A  1031;  J  1928) » 
Owner,  (for  mail),  C.  A,  Rydell  Associates,  188 
Dartmouth  St.,  Boston,  and  280  Qulnobequin 
Rd,,  Waban,  Masa, 

S 

SABIN,  Edward  R,  (M  1019).  E,  R,  Sabin  &  Co., 
4710-12  Market  St.,  Philadelphia,  Fa, 

SADLER,  C.  Boone  (M  1028),  Design  Draftsman 
(for  mall),  Public  Works  Office,  llth  Naval 
District,  and  4820  Voltaire  St.,  San  Diego,  Calif, 

SAITO,  Shoseo  (M  1923).  Marunouchl  Bldg,, 
Opposite  Tokyo  Station,  Tokyo,  Japan, 


SAKOUTA,   Mathieu  L.    (M   1924),   Consulting 

Engr.    and    Expert    Gavan,    Simanskaia    4-A, 

Leningrad,  U.  S.  S.  R, 
SANBERN,  E.  Nutc*  (Al  1023),  123  S.  Haviland 

Ave.,  Audubon,  N.  J. 
SANDS,  Clive  C.  (M  1929)   G,  P.  O.  Box  001  F.  F., 

Sydney  N.  S.  W.,  Australia. 
SANFORD,  Arthur  L.   (JU  1915),  Mech.  Engr., 

4240  Aldrich  Ave.  S.,  Minneapolis,  Minn. 
SANFORD,  Sterling  S.  (<U  1930) ,  Engr.  Sales  (for 

mail),   Detroit  Edison  Co.,  2000  Second  Ave., 

and  l.r)03  Soyburn  Ave.,  Detroit,  Mich. 
SANTEE,  Helen  G.  (At  1930),  Asst.  to  Architect 

and  Engr.,  City  School  Dist.,  81  N.  Washington 

St.,  and  (for  mail),  900  S.  Kranklin  St.,  Wilkes- 

Barre,  Pa. 
SAUER,  Robert  L.  (A  IIWO),  Dist.  Sales  Mgr.  (for 

mail),  Rilcy  Stoker  Corp.,  Kt.  of  Walker  St.,  and 

3315  W.  Philadelphia,  Detroit,  Midi. 
SAUNDERS,  Laurence  F.  (A!  19,'W),  Director  of 

Engrg.,    Harrison    Radiator    Corp.,    Lockport, 

N.  Y. 
SAWDON,  Will  M.  (M  1920),  Prof.  Experimental 

Engrg.  (for  mail),  Cornell  University,  and  1018 

E.  State  St.,  Ithaca,  N.  Y. 
SAWHILL,  R.   V.   (A    1929), _  Editor   (for  mail), 

Domestic  Engrg.,   1900  Prairie  Ave.,  Chicago, 

and  «r>34  Oakdalc  Ave.,  Glencoe,  111. 
SAWYER,  J.  Neal  (J  1933),   Production  Dept., 

Jiolland   Furnace  Co.,   and   (for  mail),  78  East 

12th  St.,  Holland,  Mich. 
SCANLON,  Edward  S.  (A  1934),  2510  Homehurst 

Ave.,  Pittsburgh,  Pa. 
SCHEIOECKER,  Dantol  B.  (A  1010),  Secy,  (for 

mail),    Hunter-Clark   Vt«.    Syatwn   Co.,   2800 

Cottage  Grove  Ave,,  ami  4020  N.  Kilbourn  Ave., 

Chicago,  III. 
SGHERNBBGK,  Fred  H.  (A  1930),  Salesman  (for 

mail),  Win.  Bros.  Boiler  &  Mf«.  Co.,  Nicollet 

Island,  and  5045  Portland  Ave.,  Minneapolis, 

Minn. 

SGHIOK,  Karl  W.  (A  1034),  Salca  Kngr.,  Minne- 
apolis-Honeywell   Regulator   Co.,   4500    Kuelid 

Ave.,  and  (for  mail),  2044  Cornell  Rd.,  Cleve- 
land, Ohio. 
SGHLICHTING,  Walter  G.  {M  1932),  Mgr.,  Air 

Cond.  Dept.,  Clara  ge  Kan  Co.,  and  (for  mail), 

1417  W.  Lovell  St.,  Kalamaxoo,  Mich. 
SCHMIDT,  Richard  H.  <5  1934),  2130  Abinston 

Rd.,  Cleveland,  Ohio. 
SGHMUTZ,  Jfoan  (M  1933),  Head  Mgr.  (for  mail), 

P,  R.  S.  M,  40  Rue  Ameiot,  Paris  Xle,  and  18 

Rue  DufrSnoy,  Paris  XVIe,  France. 
SCHNEIDER,  WJlUam  G.  (Af  15*32),  (for  mail), 

The  American  Bra$s  Co.,  25  Broadway,  New 

York,  N.  Y. 
SCHO£tfUAHN,  Robert  P,  (M  1910),  Consulting 

Engr.  (for  mail).  «*{(H-5  Industrial  Trust  Bldg., 

and  710  Nottingham  Rd.,  Wilmington,  Del, 
SGHQBNOFF,  Alfred  E.  (/  1935;  $  ItKiQ),  Secy- 

Treas.,   Schoenoff  Tlbg,  &   Htg.   Co,,   &U  E» 

Second  St,,  and  (for  mail),  513H  E,  Second  St., 

Menornonie,  Wi«. 
SCHOEPFLIN*    Paul    H,    (M    1920),    Niagara 

Blower  Co.,  0  East  45th  St.»  N«w  York,  N.  Y. 
SGHULZ,  Howard  I.  (A  1915),  Crane  Co.,  1228 

W.  Broad  St.,  Richmond,  V®. 
SGHULZE,  Benedict  H.  (M  1921).  Eastern  6ntei 

Mgr.  (for  mail),  Kewanee  Boiler  Corp.,  37  West 

39th  St.,  and  57  Purk  Ave,,  Ntw  York,  N.  Y, 
SGHUEMAN,  Jfoba  A,  (J  1035),  Air  Cond.  Enar, 

(for  mail),   York  Ic©  Machinery  Corp,,  2700 

Washington  Av«s.  N,W»,  Clev^Itnd,  and  1Q&9 

Parktfde  Dr,,  L^kewood,  Ohio, 
SGBWARTC,    Jacob     (J    1020),     Contmctor, 

Samiiel  Schwartz  8c  Son,  Inc.,  SO  west  27th  8t»« 

Bayonne.  and  (for  mail),  12  van  Hout&ia  Ave»» 

Jersey  City,  N.  J» 
SGHWEIM,  Henry  J.  (M  1828),  Chtef  Saw*  and 

Secy,  (for  mail).  Gypsum  Am.,  211  W,  WIM  " 

Dr.,  and  1012  Estes  Ave.,  Chicafo,  r 
SCOFIELP,  Paul  C,  (J  19* 

Carrier  Easw.  Corp,,  74S  ] 

and  800  N.  <Ss  ' '       "  "^J 


84 


ROLL  OF  MEMBERSHIP 


SCOTT,  Charles  E.  (M  1907),  Pres.  and  Treas. 
(tor  mail),  Vapor  Engrg.  Co.,  489  Fifth  Ave., 
New  York,  N.  Y.,  and  Darien,  Conn. 

SCOTT,  George  M.  (M  1915),  Vice-Pres.  (for 
mail),  Child  &  Scott,  Donohue,  Inc.,  112  Wooster 
St.,  New  York,  and  GO  Bowman  Ave.,  Port 
Chester,  N.  Y, 

SCOTT,  William  P.,  Jr.  (J  1933),  9  Scenic  Way, 
San  Francisco,  Calif. 

SCR1BNER,  Eugene  D.  (A  1933;  J  1929),  Engr, 
(for  mail),  Carrier  Engrg.  Corp.,  Chrysler  Bldg., 
R.  408  New  York,  N.  Y.,  and  204  Prospect  St., 
Westfield,  N.  T- 

SEEBER,  Rex  R.*  (M  1931),  Head,  Mech.  Engrg. 
Dcpt.,  Michigan  College  of  Mining  and  Techno- 
logy, Houghton,  Mich. 

SEELBACH,  Herman  (M  1931),  Pres.  (for  mail), 
Equipment  Sales,  Inc.,  610  Erie  County  Bank 
Bldu.,  Buffalo,  and  31  Central  Ave.,  Hamburg, 
N.  Y. 

SEELEY,  Lauren  E.*  (Al  1930),  Asst.  Prof,  (for 
mail),  Yale  Ensrj*.  School,  Yale  University, 
Mason  Lab.,  and  130  Everit  St.,  New  Haven, 
Conn. 

SEELIG,  Alfred  E.  (U  1920),  Pres.  and  Gen. 
Mgr,,  L.  J.  Wing  Mfg.  Co.,  154  West  14th  St., 
and  (for  mail),  310  Convent  Ave.,  New  York, 
N,  Y. 

SEELIG ,  Lester  (M  1925),  Mech.  Engr.,  Museum 
of  Science  and  Industry,  Jackson  Park,  and  (for 
mail),  725  Irving  Park  Blvd.,  Chicago,  111. 

SEBPE,  Paul  E.  (A  1933),  Sales  Engr.,  Minne- 
apolis-Honeywell Regulator  Co.,  2831  Olive  St., 
St.  Louis,  and  (for  mail),  8233  John  PI.,  Wellston, 
Mo. 

SKITER,  J.  Earl*  (M  1928),  Asst.  Mgr.,  New 
Business  Dept.,  Consolidated  Gas,  Electric 
Light  &  Power  Co.,  and  (for  mail),  7117  Bristol 
Rd.,  Baltimore,  Md. 

SEKIDO,  Kunisuke  (M  1003),  Consulting  Engr., 
Marunouchi  Bldg.,  No.  855,  and  (for  mail),  10 
Momo/Guo  Nakano,  Tokyo,  Tanan. 

SELLMA3N,  Nils  T.  (M  1022),  Director  of  Sales 
and  Utilisation,  and  Asst.  Secy,  (for  mail), 
Consolidated  Gas  Co.  of  New  York,  4  Irving 
PI.,  and  f>(5  Wnlworth  Ave.,  Seursdalc,  N.  Y. 

SENIOR,  Richard  L.  (M  1925),  (for  mail),  R.  I,, 
Senior,  Inc.,  103  Park  Ave.,  New  York,  and  10 
Cherry  Av«,»  New  Rochelle,  N.  Y. 

SENNET,  Lowell  E,  (S  1934),  1711)  East  115th  St., 
Cleveland,  Ohio. 

SBVERN8,  William  H.*  (M  1033),  Prof,  of  Mech, 
Engr,  (for  mail),  Dept,  of  Mech.  Knzr&,  Uni- 
versity of  Illinois,  and  009  Indiana  Ave.,  Urbana, 

SBWAR0,  Percival  H,*  (Charter  Member;  Life 
Member),  Research,  369  Washington  Ave., 
Brooklyn,  N.  Y, 

SHAKE,  I,  Ernest  (A  1034),  Sales  Engr.,  B.  F, 
Sturtevant  Co,,  80  Broad  St.,  Boston,  and  (for 
mail),  35  Keasenden  St.,  Dorchester,  Haas. 

SHAN  KLIN,  Arthur  P,  (M  1021)),  Sales  Engr,  (for 
mail),  Carrier  B^ngrg.  Corp.,  12  South  12th  St., 
Philadelphia,  and  40  Amherst  Ave.,  Swarthmore, 
Pa, 

SHANKXINt  JoHn  A.  U4  1028),  Secy-Treas,  (for 
mail).  West  Virginia  Hfcg.  &  Plbg,  Co,,  283  Hale 
St.,  and  1507  Quarrier  St.,  Charleston,  W.  Va. 

SHARP,  Floyd  H.  (M  1920-),  H7  E,  Third  St., 
Jamestown,  N.  Y» 

SHARP*  Htmrv  C,  (M  1935),  Mgr,,  Oil  Heat  DIv, 
-(for  mafl),  Smith  Oil  &  Reining  Co.,  HQ£ 
Kilburn  Ave,,  and  1928  Rockton  Ave.r  Rock- 
ford.  III. 

SHAVER,  Herbert  H,  (A  1*39).  Asat.  Gen.  Sales 
Agent  (for  mail),  Hudton  Coal  Co.,  424  Wyoming 
Ave.,  and  1507  Wyoming  Ave.t  ScrtntGn,  Pa, 

SHAW,  Burton  E.  (/  1034),  Research  Chief, 
Ollbtrt  &  Barker  Mfo  Co.,  Springfield,  and  (for 
mail),  Gnmby  Rd,,  Southwlck,  Mass. 

SHAW,  Mftar  (M  1920),  Prea.  (for  mail),  tynch 
&  Woodward,  lac,,  820  Dow  St.,  Bottom  and 
WT&oyal  St,  Woilaeton,  Mass, 
W,  Harold  W«a<m  t$  1905), 
«M  &t  Paul, 


SHAW,  Norman  J.   H.    (M  1927;  J  1925),  37 

Benjamin  Rd.,  Arlington,  Mass. 
SHAWUN,  Walter  C.  (A  1931),  696  S.  Oak  Park 

Court,  Milwaukee,  Wis. 
SHEA,  Michael  B.   (M  1921),  Sales  Dept.   (for 

mail),  American  Radiator  Co.,  1344  Broadway, 

Detroit,  and  114  Massachusetts  Ave.,  Highland 

Park,  Mich. 
SHEARS,   Matthew  W.    (U  1922),   39   Sylvan 

Ave.,  Toronto,  Ont.,  Canada. 
SHEFFLER,  Morris  (M  1921),  Pres,  (for  mail), 

Sheffler-Gross  Co.,  203  Drexel  Bldg.,  and  5451 

Lebanon  Ave.,  Philadelphia,  Pa. 
SHELDON,  Nelson  E.  (M  1927),  Dist.  Sales  Mgr. 

(for  mail),   Carrier  Engrg.   Corp.,  916  Temple 

Bldg.,  and  41  Lanark  Crescent,  Rochester,  N.  Y. 
SHELDON,    William   D.,   Jr.    (J    1934),    Chief 

Engr.,  Sheldons,  Ltd.,  and  (for  mail),  Cedar  St., 

Gait,  Ont.,  Canada. 
SHELNEY,  Thomas  (M  1931),  Pres.  (for  mail), 

Pierce  Blower  Corp.,  27  Carolina  St.,  and  Hotel 

Fillmore,  Buffalo,  N.  Y. 
SHENK,   Donald  Hugh   (M  19341,   10G  Forest 

Lane,    and    (for    mail),    Riggs    Hall,   Clemson 

College,  S,  C. 
SHEPARD,   Edward  C.    (M  1932),   Owner   (for 

mail),  Shepard  Engrg.  Co.,  370  Lexington  Ave., 

and  978  Grant  Ave,,  New  York,  N.  Y. 
SHEPARD,  John  deB.    (J   1929),   Consolidated 

Gas,  Electric  Light  &  Power  Co.,   Room  406 

Lexington  Bldg.,  Baltimore,  Md. 
SHEPPARD,  Frank  A.  (M  1918),  Salesman  (for 

mail),  Johnson  Service  Co.,  411  East  10th  St., 

and  27  East  70th  St.,  Kansas  City,  Mo. 
SHEPPARD,  William  G.  F.  (M  1922),  Partner 

(for  mail),   Shcppard   &  Abbott,   119   Harbord 

St.,    and    1    Clarendon    Ave.,    Toronto,    Ont., 

Canada. 
SHERET,  Andrew  (M  1929;  A  1925),  Pres.  (for 

mail),  Andrew  Sheret,  Ltd.,  1114  Blanshard  St., 

and  1030  St.  Charles  St.,  Victoria,  B.  C.,  Canada. 
SHERMAN,  Ralph  A.  (M  1933),  Fuel  Engr.  (for 

mail),    Buttolle   Memorial    Institute,   505   King 

Ave.,  and  1893  Coventry  Rd.,  Columbus,  Ohio. 
SHIVERS,  Paul  F.  (M  1930),  Chief  Engr.,  Minne- 
apolis-Honeywell Regulator  Co.,  Wabash,  Ind. 
SHODRON,  John  G.  (M  1921),  Consulting  Engr. 

and    Research,   419    E.    Milwaukee   Ave.,    Ft. 

Atkinson,  Wis. 
SHGRB,  Will  A.  (M  1909),  Treas.,  The  Field  & 

Shorb  Co.,   705  N.  Pine  St.,  and   (for  mail), 

3  Lincoln  PL,  JDecatur,  111. 
SHROCK,  John  H.  (M  1924),  Mgr.  (for  mail), 

New  York,  Blower  Co.,  Factory  St.,  and  1524 

Michigan  Ave.,  La  Porte,  Ind. 
SHULTZ,  Earl©  (A  1919),  Vice-Prea.  (for  mail), 

Illinois  Maintenance  Co,,  1136-72  W.  Adams"St., 

and  Edgewater  Beach  Apts.,  Chicago,  111. 
SIEB&   Claude   T,    (A   1927),   Service  Systems 

Kngr.   (for  mail),   western   Electric  Co,,   Inc., 

195  Broadway,  New  York,  N.  Y.,  and  Russell 

Rd.,  Fanwood,  N.  J. 
SIECEL,  L«o  (M  1928;  A  1925),  Mech,  Engr., 

1016  Lancaster  Ave.,  Brooklyn,  N.  Y* 
SIGMUND,  Ralph  W.  (M  1932),  Dist,  Mgr.  (for 

mail),  B.  F.  Sturtevatit  Co.,  913  Provident  Bank 

Bldg.,  and  S04  Oak  St.,  Cincinnati,  Ohio. 
SIMKIN,  Milton  (J  1935;  S  1938),  103  Brighton 

Ave,,  Perth  Amboy,  N.  J. 
SIMQNDS»  Abe  H.  (A  1935;  J  1929),  Sales  Engr. 

(for  mail),  Carrier  Engrg.  Corp.  of  California,  74,8 

E,  Washington  Blvd.,  and  440  Westminster  Ave,,  * 

Loa  Angeles,  Calif. 

SIMPSON,  Donald  C.  (M  1932),  Supt.  of  Re- 
search, Industrial  Mat.  Biv,  (for  mail),  Owena, 

Illinois  Glass  Co.,  Newark,  ana  878  JCelton  Ave,, 

Columbus,  Ohio. 
SIMPSON,  William  KU  (M  1919),  Vice-Pres.  (for 

mail),  Hoffman  Specialty  Co.,  an4  0  Sands  St., 

Watertmry.  Conn. 
SK,H>MORB?  Jfohti  G,  (J  19SO),  Air  Cond.  Engr., 

Carrier  Engrg.  Corp,,  408  Chrysler  Bldg.,  New 

York,  and   (lor  mail),  5101*a9th   Ave.,  tong 

Island  City,  N<  Y, 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


SKINNER,  Henry  W.  (M  1920),  Consulting  Engr. 

(for   mail),    Box   1334,   and   4816   Dexter,    Ft. 

Worth,  Texas. 
SKXENARIK,  Louis  (J  1928),  305  East  72nd  St., 

New  York,  N.  Y. 
SLAYTER,  Games  (M  1931),  OS  Walhalla  Rd., 

and  (for  mail),  711  Southwood  Ave.,  Columbus, 

Ohio. 

SLIGHT,  Irvin  (A  1925),  Slight  Bros.,  741  York- 
way  PL,  Jenkintown,  Pa. 
SMAK,   Julius  R.    (A    1934),   Supt.   of  Service 

Depts.,  Crane  Co.,  South  Ave.,  and  (for  mail), 

3135  Park  Ave.,  Bridgeport,  Conn. 
SMALL,  Bartlett  R.  (J  1932),  Sales  Engr,  (for 

mail),  T.  C.  Heyward,  1408  Independence  Bldg., 

and  326  West  10th  St.,  Charlotte,  N.  C. 
SMALL,  John  D.*  (M  1910),  Consulting  Engr. 

(for  mail),  127  N,  Dearborn  St.,  Chicago,  and 

411  Maple  Ave.,  Wilmette,  111. 
SMALLMAN,  Edwin  W.  (M  1920),  Navy  Dept., 

and  (for  mail),  S3J  Allison  St.  N.W.,  Washington, 

D.  C. 
SMITH,   Elmer  G.*   (M  1929),  Asst.   Prof,   of 

Physics,  Agricultural  and  Mechanical  College  of 

Texas,  College  Station,  Texas. 
SMITH,  Card  W.  (M  1927),  Salesman,  Premier 

Warm  Air  Heater  Co,,  Dowagiac,  Mich.,  and  (for 

mail),  248  Roche  St.,  Huntington,  Ind. 
SMITH,  Jared  A.  (A  1933),  Br.  Mgr.  jffor  mail), 

The  Bryant  Heater  &  Mfg.  Co.,  626  Broadway, 

and  3817  Indian  View  Ave.,  Mariemont,  Cin- 
cinnati, Ohio. 
SMITH,  J.  Darrell  (M  1933),  Mech.  Engrg.  Dept, 

Philadelphia  Si  Reading  Coal  &  Iron  Co.,  and  (for 

mail),  317  North  19th  St.,  Pottsville,  Pa. 
SMITH,   Milton   S.    (M  1019),   Trcas.,   Carrier 

Engrg.  Corp.,  850  Frelinghuysen  Ave.,  Newark, 

and  (for  mail),  13  N.  Terrace,  Maplewood,  N.  J. 
SMITH,  Robert  Hugh  (7  1934;  S  1933),  Sales 

Engr.,  Sears    Roebuck  &  Co,,  Dept.  of  Plbg., 

Htg.  and  Vtg.,  135  S.  Fifth  St.,  and  (for  mail), 

214  N.   Fourth   St.,   Room  410,   Steubenville, 

Ohio. 
SMITH,  Wilbur  F.  (M  1920),  Consulting  Engr., 

000  Schuylkill  Ave.,  Philadelphia,  and  (for  mail), 

422  Bryn  Mawr  Ave.,  Cynwyd,  Pa. 
SMOOT,  Thco  Halley  (M  1935),  Chief  Engr,  (for 

mail),  Fluid  Heat  Div,,  Anchor  Post  Fence  Co., 

Eastern  Ave.  and  Kane  St.,  and  2512  Talbot  Rd., 

Baltimore,  Md. 
SHYERS,   Edward  €.    (A    1933),   Sales  Engr., 

Minneapolis-Honeywell    Regulator    Co.,    1013 

Penn  Ave.,  Wilkinaburg,  and   (for  mail),   148 

Jamaica  Ave.,  West  View,  Pittsburgh,  Pa. 
SNEED,  Richard  B.  (S  1934)>  (for  mail),  College 

of  Engineering,  University  of  Oklahoma,  Norman 

and  Brlstow,  Okla, 
SHELL,  Ernest  (M  1920)  ,'39 14  LeMay  Ave,, 

Detroit,  Mich. 
SNIDER,  Lewis  A.  (M  1927),  Pres.  (for  mail), 

L.  A,  Snirler  Engrg.  Service,  Inc.,  005  N.  Michi- 
gan Ave. /and  049  Buena  Ave.,  Chicago,  111. 
SNYDER,  Allen,  K.  (J  1930),  Air  Cond,  Engr., 

Richmond  Air  Equipment  Co.,  Inc.,  1804  W. 

Broad  St,  and  (for  mail),  4309  Grove  Ave., 

Richmond,  Va. 
SNYDER,   Jay  W.    (M   1017),   McColl-Snyder- 

McLean,  2304  Pcnobscot  Bldg,,  Detroit,  Mich, 
SNYDER,   Joseph   S,    (A    1025),   Sales   Repr., 

Detroit  Lubricator  Co.,  374  Delaware  Ave.,  and 

(for  mail),  9  Knowlton  Ave,.  Buffalo,  N,  Y. 
SODEMANN,  Paul  W.  (M  1«;  J  1920),  Sales 

Engr.,  2300  Delraar  Blvd.*  and  (for  mail),  4130 

Parlln  Ave,,  St.  Loute,  Mo. 
SODEMANN,  William  C.  B.  (M  1019),  Pre»,  (for 

mail),    Sodenmnn    Heat   &    Power   Co,,    2308 

Delmar  Blvd.,  St.  Louis,  Mo, 
SONNEBORN,  Gharkss  (M  1930),  Vice-Pres.  in 

charge  of  Production,  Shaw,  Perkins  Mfg.  Co., 

West  Pittsburgh,  and  (for  mail),  R.  D.  No.  3, 

New  Castle,  Pa. 
SONNEY,  Kermit  J,  (S  1034),  (for  mail),  310 

W,  Symmea  St.,  Norman,  0kla,«  and  L,  B,  lad, 

WIlcox,  Pa. 
SOPER.  H.  A»  (U  1910),  Vice-Free.,  American 

Foundry  &  Furnace  Co,,  Bloomlngton,  HI, 


SOULE,  Lawrence  C.*  (M  1908),  Secy,  and  Chief 
Engr.  (for  mail),  Aerofin  Corp.,  850  Freling- 
huysen Ave.,  Newark,  and  Essex  Fells,  N.  J. 

SPAFFORD,  Allen  (A  1027),  Wood  Conversion 
Co.,  Cloquet,  Minn. 

SPECKMAN,  Charles  H.  (M  1918),  Prof.  Engr., 
375  Bourse  Bldg.,  Philadelphia,  Pa. 

SPFXLER,  Frank  N.*  (M  190S),  Director,  Dept. 
of  Metallurgy  and  Research  (for  mail).  National 
Tube  Co.,  1922  Frick  Bldg.,  and  C-ill  Darlington 
Rd.,  Pittsburgh,  Pa. 

SPENCE,  Morton  R.  (J  1934),  Asat.  Purchasing 
Agent,  Rundle  &  Spence  Mf£.  Co.,  445  N. 
Fourth  St.,  and  (for  mail),  709  E.  Lexington 
Blvd.,  Milwaukee,  Wis. 

SPENCER,  Roland  M.  (J  1934),  Sales  Enpjr,  (for 
mail),  Powers  Regulator  Co.,  754  Hippodrome 
Annex,  Cleveland,  and  1269  Bonnie  View  Ave., 
Lakewood,  Ohio. 

SPIELMAN,  Gordon  P.  (A  1031;  J  1923), 
Harrison,  Spielman  Co.,  480  Milwaukee  Ave., 
Chicago,  111. 

SPIELMANN,  Harold  J.  (M  1933),  Air  Cond. 
Engr.,  The  Vilter  Mfg.  Co.,  and  (for  mail),  2549 
N.  Lake  Dr.,  Milwaukee,  Wis. 

SPITZLEY,  Ray  L.  (M  1920),  1200  W.  Fort  St., 
Detroit,  Mich. 

SPOFFORTH,  Walter  (M  1930),  Chief  of  Mech. 
Services,  U.  S.  Penitentiary,  McNeil  Island,  and 
(for  mail),  1850  W.  Blvd.,  Day  Island,  Tacoma, 
Wash. 

SPROULL,  Howard  B.  (M  1920),  Div,  Sales  Mgr. 
(for  mail),  American  Blower  Corp.,  1005-H 
American  Bids-,  and  3588  Ray  mar  Dr.,  Cin- 
cinnati, Ohio. 

SPURGEON,  Joseph  H.  (M  1924),  Salesman  (for 
mail),  Spurgeon  Co.,  5-203  General  Motors 
Bld#.,  and  17215  Pcnnington  Dr.,  Detroit,  Mich. 

STAGEY,  Alfred  E.,  Jr.*  (M  1914),  Wooton  Rd,, 
Essex  Fells,  N,  J. 

STACK,  Frank  Charles  (J  1035;  S  1933),  140-23 
Cherry  Ave.,  Flushing  N.  Y. 

STACY,  Stanley  C.  (M  1931),  Mech*  Engr.  (for 
mail),  Board  of  Education,  13  S,  FiUhugh  St., 
and  91  Cobfos  Hill  Dr.,  Rochester,  N.  Y. 

STALB,  Joseph  O.  (A  1934),  New  York  Mgr., 
Parco  Furnace  Div,,  Reading  Iron  Co.,  143 
Liberty  St.,  and  (for  mail),  22  Partridge  Ave., 
Ridley  Park,  Pa. 

STAMMER,  Edward  L.  (M  1919),  Supt.,  Htg. 
and  Vtg.,  Board  of  Education  Bldg.,  and  (for 
mail),  4430  Tennesee  Ave.,  St.  Louis,  Mo. 

ST  ANGER,  Ralph  B.  (M  X920),  Owner  (for mail), 
Robinson  &  Stanger,  Empire  IJldg,,  Pittsburgh, 
and  Deer  Creek,  Church  Rd.,  Glenshaw,  Pa, 

STANOLAND,  &.  F,  (Charter  Member),  (2nd 
Vice-Pres,,  1908;  Bourd  of  Governors,  1005, 
1906,  1909;  Board  of  Mgrs.»  1895-1899;  Council, 
1800- 18»7),  Kendall,  N.  Y. 

STANNARD,  James  M.*  (Life  Mm&flv  M  1906), 
Prea-Treas,  (for  mnil),  StannartI  Power  Equip- 
ment Co.,  £3  W,  Jackson  Klvd,,  Chicago,  and 
1403  KHnor  PI.,  Kvanston,  fit, 

STAPLES,  William  II.  (A  1924),  Htg,  and  Vt«. 
(for  mail),  Maguire  Staples  tfe  Mason,  24  West 
20th  St.,  and  £48  West  1 64th  St.,  New  York, 
N.  Y. 

STARK,  W,  Elliott*  (M  1020),  (Council,  1932- 
1934.),  Research  Engr,,  Bryant  Iteiter  &  Mfg, 
Co,,  17825  vSt.  Glair  Ave,.  Cleveland,  and 
(for  mail)*  1875  Rosemontt  Rc3U  East  Cleveland, 
Ohio, 

STEELS,  John  B.  (M  1932),  Chief  Engr.  (for 
mail).  Engrg.  Dept,*  Winnipeg  School  Board, 
Ellen  and  William  Avev,  and  184  Waterloo  St, 
Winnipeg,  Man,,  Canada, 

STEELEf  Maurice  G.  (M  102%  Product  Engr. 
(for  mail).  Revere  Copper  &  Br&ts,,  Inc»»  Re- 
search Dept,,  and  006  N.  Mudlson  St*,  Rome, 
N.  Y. 

STEEN,  Joseph  M*  (M  1929),  Iron  City  Kt&  Co,, 
843  Jaeksonia  St.,  Pittsburgh,  Pa. 

STEFFNER,  Edward  F.  (J  1934),  1Q&17  Fortuae 
cU  Ohio, 


36 


ROLL  OF  MEMBERSHIP 


STEGGALL,  Howard  B.  (A  1934),  Br.  Mgr.  (for 

mail),  U.  S.  Radiator  Corp.,  941  Behan  St.,  and 

1166  Murray  Hill  Ave.,  Pittsburgh,  Pa. 
STEINHORST,  Theodore  F.   (M  1919),  Treas. 

and  Gen.  Mgr.,  Emil  Steinhorst  &  Sons,  Inc., 

612  South  St.,  and  (for  mail),  1664  Brinckerhoff 

Ave.,  Utica,  N.  Y. 
STEINKELLNER,  Edward  J.  (S  1935),  315-19th 

Ave.  S.E.,  Minneapolis,  Minn. 
STEINMETZ,  C.  W.  Arthur  (M  1934),  Mgr.  (for 

mail),  American  Blower  Corp.,  402  Broad  St., 

Newark,  and  50  Oakwood  Ave.,  Bogota,  N.  J. 
STEPHENSON,  L.  A.  (M  1917),  Mgr.  (for  mail), 

Powers  Regulator  Co.,  409  East  13th  St.,  and 

801  West  57th  Terrace,  Kansas  City,  Mo. 
STERNBERG,    Edwin    (A    1932;    J   1931),    Air 

Cond.  Engr.,  Arctic  Engrg.  Co.,  123  White  St., 

and  (for  mail),  58  East  92nd  St.,  New  York,  N.  Y. 
STERNE,  Cecil  M.  (A  1934),  Chief  Engr.   (for 

mail),  Metropolitan  Refining  Co.,    Inc.,    23-28 

50th  Ave,,  Long  Island  City,  and  115  Harold 

Rd.,  Woodmere,  L.  L,  N.  Y. 
STETSON,  Lawrence  R.  (M  1913),  303  Congress 

St.,  Boston,  Mass. 
STEVENS,  Harry  L,  (M  1934;  A  1927;  J  1924), 

Secy-Treas.  (for  mail),  M.  M.  Stevens  Co.,  108 

W.  Sherman,  and  7  West  22nd  St.,  Hutchinson, 

Kans. 
STEVENS,  John  M.  (A  1933),  4643  Morris  St., 

Philadelphia,  Pa. 
STEVENS,  William  R.  (A  1934),  Partner,  L.  B. 

Stevens  Co.,  442  E.  Front  St.,  Cincinnati,  Ohio, 

and  (for  mail),  159  Tremont  Ave.,  Ft.  Thomas, 

Ky. 
STEVENSON,  Wilbur  W.  (M  1928),  Steam  Htg. 

Engr*  (for  mail),  Allegheny  County  Steam  Htg. 

Co*,  435  Sixth  Ave.,  and  1125  Lancaster  Ave., 

Pittsburgh,  Pa, 
STEWART,    Charles   W.    (M    1919;    A    1918), 

Asst.  Secy,  (for  mail),  Hoffman  Specialty  Co., 

and  21  Yates  Ave.,  Watcrbury,  Conn. 
STEWART,  Clement  W.  (M  1934),  Sales  Engr, 

(for  mail),  Ilg  Electric  Vtg,  Co.,  15  Park  Row., 

and  3985  Saxon  Ave.,  New  York,  N.  Y. 
STEWART,  Duncan  J.  (A  1930),  Mgr.,  Electric 

Apparatus  Div.  (for  mail),  Barber-Colman  Co., 

and  214  Franklin  PI,,  Rockford,  111. 
STEWART,  John  C.  (A  1934),  Owner  (for  mail), 

1844  Smith  St.,  and  2807  Victoria  Ave.,  Regina 

Saek,  Canada. 
STILL*  Fred  R,*  (M  1004),  (Presidential  Member), 

(Prea,,    1918;   2nd   Vicc-Pres.,    1017;    Council, 

1916-1919),    Vice-Pres,    (for    mail),    American 

Blower  Corp.,  401  Broadway,  and  1  East  End 

Ave.,  New  York,  N.  Y. 
STILLER,  Frederick  Wilbur  (J  1933),  Estimator 

(for  mail),  F.  C.  Stiller  &  Co.,  121)  S.  Tenth  St., 

and  138  West  49th  St.,  Minneapolis,  Minn. 
STINARD,    Rutherford    L.    (J    1934),    Engr., 

American  Radiator  CoM  40  West  40th  St.,  New 

York,  N.  Y,.  and  (for  mail),  1377  Boulevard  B, 

West  New  York,  N,  J. 
STITT,  Arthur  B,  (J  1085;  S  1933),  Plbg,  and 

Htg,  Bngr,»  Sears  Roebuck  &  Co.,  184  Atlantic 

St.,  Stamford,  Conn.,  and  (for  mail),  260  Valen- 

tine Lane*  Yonkers,  N,  Y. 
STKTT,  Eugenia  W*  (M  1917),  Sales  Repr.,  Cast 

Product®    Div.    (for   maii),   46   Second   Ave», 

Johnstown,  Pa, 
STOCK.WBLL,  William  R.  (M  1903;  J  1901), 

Weil-McLain  Co,,  Michigan  City,  IndU 
STONE,  Euftene  R.  (M  1913),  78  Woodbine  St., 
y,  Mass. 


STONE,  Georlfc  F,  (Life  Member;  M  1918), 
Estimator,  19  Blmwood  Rd.»  Verona,  N.  J. 

STRAUCH,  Paul  C.  (A  1984),  Sales  Engn,  The 
Henry  Furnace  &  Foundry  Co.,  18th  and 
Merriman  Sta,,  Pittsburgh*  and  (for  mall),  101 
Washington  Ave,,  Edgewood,  Pittsburgh,  Pa, 

STRBVBLL,  Rofter  P.  (M  19340,  Co-Partner  (for 
mall),  Wra*  R,  Hogg  Co.,  900  Fourth  Ave,» 
Asbury  Park,  and  corner  State  Highway  and 
Victor  R,  Neptune,  N,  J. 


STRICKLAND,  Albert  W.  (A  1929),  Htg.  and 

Vtg.  Engr.,  Big  Timber,  Mont. 
STROCK,  Clifford  (A  1929),  Associate  Editor  (for 

mail),  Heating  and  Ventilating,  148  Lafayette 

St.,  and  150  East  182nd  St.,  New  York,  N  Y. 
STROUSE,  Sherman  W.  (A  1934),  Sales  Engr., 

Cooney  Refrigeration  Co.,  Inc.,  and  (for  mail) 

315  Capen  Blvd.,  Buffalo,  N.  Y. 
STROUSE,  Sidney  B.  (M  1921),  Engr.  (for  mail) 

500-529    Guarantee    Trust    Bldg.,    and    22    S. 

Illinois  Ave.,  Atlantic  City,  N.  J. 
STRUNIN,  Jay  (J  1933),  Engr.  and  Contractor 

(for  mail),  Strunin  Plbg.  &  Htg.  Co.,  408  Second 

Ave.,  and  54  West  89th  St.,  New  York,  N.  Y. 
STUBBS,  W.  C.   (M  1934),  Design  Draftsman, 

Heat  and  Vtg.  (for  mail),  Norfolk  Navy  Yard, 

and  36  Chanmng  Ave.,  Portsmouth,  Va. 
SUMMERS,  Ernest  T.  (A  1930),  Pres.  (for  mail), 

Summers,  Darling  &  Co.,  121  Smith  St.,  and  Ste. 

22  Newcastle  Apts.,  Winnipeg,  Man.,  Canada. 
SUNDELL,  Samuel  S.   (J  1935;  5  1933),  3040 

Longfellow  Ave.  S.,  Minneapolis,  Minn. 
SUPPLE,  Graeme  B.  (M  1934),  American  Blower 

Corp.,    025    Architects    and     Builders    Bldg., 

Indianapolis,  Ind. 
SUTCLIFFE,  Arthur  G.  (M  1922;  A  1918),  Chief 

Engr.,  Ilg  Electric  Vtg.  Co.,  2850  N.  Crawford 

Ave.,  and  (for  mail),  4146  N.  St.  Louis  Ave., 

Chicago,  111. 
SUTHERLAND,  David  L.  (A  1934),  Pres-Treas. 

(for  mail),   Sutherland   Air   Cond.   Corp.,   627 

Marquette    Ave.,    and    1815    Colfax    Ave.    S., 

Minneapolis,  Minn. 
SUTTON,  Frank  (M  1932),  Consulting  Engr.  (for 

mail),  140  Cedar  St.,  New  York,  and  Babylon, 

L.  L,  N.  Y. 
SWANEY,  Carroll  R.  (M  1929;  /  1921),  Gilbert 

Howe   Gleason,   25   Huntington   Ave.,    Boston, 

Mass. 
SWANSON,  Harry  (M  1933),  Engr.  (for  mail), 

The  Fels  Co.,  42  Union  St.,  Portland,  and  Box 

135,  Cape  Cottage,  Maine. 
SWANSON,  Rolf  G.  (S  1935),  324  Walnut  St. 

S.E.,  Minneapolis,  Minn. 
SWANSTROM,    Alfred   E.    (J   1935;   S   1932), 

Construction  Foreman,  U.  S.  Dept,  of  Interior, 

and  (for  mail),  1444  Van  Buren  St.,  St.  Paul, 

Minn. 
SWEATT,  Charles  H.  (S  1935),  4259  Unity  Ave., 

Robbinsdale,  Minn. 
SWEIVEN,  C.  E,  (S  1935),  406  Walnut  St.  S.E., 

Minneapolis,  Minn. 
SWENSON,  John  E»  (A  1930),  Industrial  Engr. 

(for  mail),    Minneapolis   Gas   Light   Co.,   800 

Hennepin  Ave.,  and  1102  South  East  13th  Ave., 

Minneapolis,  Minn. 
SWISHER,  Stephen  G.,  Jr,  (A  1934),  Sales  Engr. 

(for  mail),  The  Trane  Co.,  125  E.  Wells  St.,  and 

4238  N.  Woodburn  Ave.,  Milwaukee,  Wis. 
SYSKA,  Adolph  G.  (M  1933),  Consulting  Engr., 

Syska  &  Hennessy,  420  Lexington  Ave.,  New 

York,  N.  Y, 
SZEKEtY,  Emeat  (M  1920),  Vice-Prea.  and  Gen, 

Mgr.  (for  mail),  Bayley  Blower  Co.,  1817  South 

60th  St.,  and  3104  WJKilbourn  Ave.,  Milwaukee, 

Wis. 
S2OMBATHY,   Louis  R.    (A    1930),   Ferguson 

Sheet  Metal  Works,  Inc.,  34  N,  Florissant  Blvd., 

Ferguson,  Mo. 


TABOR,  Charles  B*  (S  1935),  1815  University 
Ave,  S.E.,  Minneapolis,  Minn, 

TAGGART,  Ralph  C,*  (M  19*12),  14  Lyon  Ave., 
Menands,  Albany,  N.  Y. 

TAUAFERRO,  Robert  R.*  (M  1919),  Air  Cond, 
Engr,,  Philadelphia  Saving  Fund  Society,  12 
South  12th  St,  Philadelphia,  and  (for  mail), 
$38  Beechwood  Rd.,  Upper  Darby,  Pa. 

TALLMADGE,  Webster  (M  1924),  Prea.  (for 
mail),  Webster  Tallmadge  &  Co,,  Inc.,  255  North 
18th  St,,  East  Orange,  and  7  Clareraont  PL, 
Montclair,  N.  J, 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


TARR,  .Harold   M.    (M  1931),    Htg.   Engr.,   21 

Montague  St.,  Arlington  Heights,  Mass. 
TAUSON,  Peter  O,  (S  1934),  408  Southwest  23rd 

St.,  Oklahoma  City,  Okla. 

TAVANLAR,  Eligio  J.*  (J  1931),  (for  mail), 
Carrier  Research  Corp.,  850  Frelinghuysen  Ave., 
Newark,  N.  J.,  and  Binalonan  Pangasinan, 
Philippine  Islands. 

TAVERNA,  Frederick  F.  (M  1928;  A  1927; 
J  1924),  Engr,,  Raisler  Htg.  Co,,  129  Amsterdam 
Ave.,  New  York,  N.  Y.,  and  (for  mail),  40G-12th 
St.,  Union  City,  N".  J. 

TAYLOR,  Edward  M.  (A  1934),  Draughtsman 
and  Engrg.  Asst.,  City  Engineer's  Dept.,  and  (for 
mail),  102  Innes  Rd.,  Christchurch,  New  Zealand. 
TAYLOR,  William  E.  (A  1934),  Factory  Sales 
Engr.,  Air  Cond.  Div.  of  (for  mail),  Gar  Wood 
Industries,  Inc.,  7924  Riopelle  St.,  and  200S  W. 
Grand  Blvd.,  Detroit,  Mich. 

TAZE,  Donovan  L.  (M  1031),  Sales  En«r., 
American  Blower  Corp.,  601)0  Russell  St., 
Detroit,  Mich. 

TEASDALE,  Lawrence  A.  (M  1920),  Partner  (for 
mail),  Oftice  of  Hollis  French,  20  Ashmun  St., 
and  262  West  Rock  Ave.,  New  Haven,  Conn. 
TEELING,    George    A.    (M    1930),    Consulting 

Engr.,  UN.  Pearl  St.,  Albany,  N.  Y. 
TEMPLE,   Walter   J.    (M    11TO,    Engr,,   J.   A. 
Temple  &  Co.,  919  E.  Michigan  Ave.,  and  (for 
mail),  1215  Reed  St.,  Kalamazoo.  Mich. 
TEMPLIN,  Charles  L.  (M  1921),  Sales  Engr.  (for 
mail),  Carrier  Engrg.  Corp.,  Bona  Allen  Bldg., 
and  781  Sherwood  Rd.  N.K.,  Atlanta,  Ga. 
TENKONOHY,  Rudolph  J.  (A/  1923),  Vice-Pres. 
(for  mail),  Airtherm  Mfg.  Co.,  1474  S.  Vande- 
venter  Ave.,  St.  Louis,  Mo,,  and  5019  Ridgewood 
Ave,  Detroit,  Mich. 

TENNANT,  Raymond  J.  J.  (.4  1929),  Supervisor 
of  Sales  (for  mail),  Duqucsne  Light  Co.,  -135 
Sixth  Ave,,  and  529  Navato  PL,  Pittsburgh,  Pa. 
TENNEY,  Dwight  (M  1032),  Pres.  and  Chief 
Engr.  (for  mail),  Tenney  Engrg.,  Inc.,  Bloomfield, 
Ave.,  at  Grove  St.,  Bloomtiekl,  and  33  Summit 
Rd.,  Verona,  N.  J. 

THEORELL,  Hugo  G.  T.*  (Life  Member;  M 
1902),  Consulting  Engr.,  Hugo  Theorells  In- 
genieussbyra,  Skoldungagatan  4,  Stockholm, 
Sweden. 

THINN,  Christian  A.*  (M  1921),  Chief  Engr., 
C.  A.  Dunham  Co.,  450  K.  Ohio  St.,  Chicago,  111. 
THOMAS,  L.  G.  Lee  (M  1934),  Vice-Pres,  (for 
mail),  Economy  Pumping  Machinery  Co.,  3431 
Weat  48th  PI.,  Chicago,  and  426  Forest  Ave., 
Oak  Park,  III. 

THOMAS,  Melvern  F,  (M  1900),  Consulting 
Engr,  (for  mail),  Thomas  &  Ward  ell,  W)  College 
St.,  and  24  Rivercrest  Rd.,  Toronto,  Gnt, 
Canada, 

THOMAS,  Norman  A.  (M  19^8),  Pres.  (fur  mail), 
Thomas  Htg,  Co.,  llth  and  Herrick  Ave.,  and 
824  Monroe  Ave.,  Racine,  Wis. 
THOMAS,  Richard  H.  (Life  Member,  M  192Q), 
Economy  Pumping  Machinery  Co,,  3431  West 
48th  PI.,  Chicago,  111. 

THQMMEN,    Adolph    A,    (A    1929),    Forman, 

Bloomer  Htg,  &  Vtg,  Co,r  1245  West  47th  St., 

and  (for  mail),  3400  West  Olst  PI.,  Chicago,  111, 

THOMPSON,  Donald  (J'  1088),  Engrg,  Dept,, 

Carbide  &  Carbon  Chemicals  Corp.,  and  (for 

mail),  514  Simma  St.,  Charleston,  W,  Va. 

THOMPSON,  Nelson  S.*  (M  1917;  J  1807),  1615 

Hobart  St.  N.W.,  Washington,  D.  C, 
THOMSON.  Thomas  N.*   (M1899)   Consulting 

Engr,,  87  Irwin  PL,  Huntinjfton,  L,  L,  N,  Y* 
THORNBVRG,  Harold  A»   (M  1932;  A   1982; 
J  1920),  Sales  Engr.  (for  mail).  Carrier  Enarg. 
Corp,,  U  South  12th  St.,  and  2115  Chestnut  St., 
Philadelphia,  Pa, 
THORNTON,  Roger  T.  (M  1910),  Buffalo  Forge 

Co.,  490  Broadway,  Buffalo,  N.  Y. 
THORNTON.  William  B.*  (M  1931),  Sales  Bngr. 
(for  iaall}(  Carrier  Ensrg*  Corp,,  404  Bona  Allen 
Bldg.,  Atlanta,  a»d  155  Coventry  Rdn  l>eatur, 
Ga. 


THRUSH,  Homer  A.  (If  1918),  H.  A.  Thrush  & 

Co.,  21-23  E.  Riverside  Dr.,  Peru,  Ind. 
TIBBETS,    John   G.    (M    1920),    Engrg.    Dept., 
B.  &  O.  R.  R.  Co.,  and  (for  mail),  P.O.  Box  106, 

Ellicott  City,  Mel 
TILLER,  Louin  (J  1935;  5  1933),  1724  Northwest 

20th  St.,  Oklahoma  City,  Okla. 
TILTZ,  Bernard  E.  (M  1930),  Pres.  (for  mail), 

Tiltz  Air  Cond.  Corp.,  285  Madison  Ave.,  New 

York,  and  24  Barnum  Rd.,  Larchmont,  N,  Y. 
TIMMIS,  Pierce  (Af  1920),  Service  Equip.  Dept. 

(for   mail),    United   Engineers   &   Constructors, 

Inc.,  1401  Arch  St.,  Philadelphia,  and  202  Mid- 

land Ave.,  Wayne,  Pa, 
TIMMIS,  W.  Walter  (M  1933;  A  1025),  Engr.  (for 

mail),  American  Radiator  Co.,  40  West  40th  St., 

New  York,  and  32  Oak  Lane,  Glen  Cove.,  N.  Y. 
TISNOWER,  William  (JU  1923),  131  Livingston 

St.,  Brooklyn,  N.  Y. 
TITUS,  Marvin  S.   (M  1028),  414  Fayctte  St 

Charleston,  W.  Va. 
TJERSLAND,  Alf  (M  1910;  J  1900),  K.  Sunde  & 

Co.,  Ltd.,  Oslo,  Norway. 
TOBIN,  George  J.   (M  190o),  Owner,  vSanitary, 

Htg.  &  Vtg.  Engr.,  1S7  North  Ave.,  Plainiield, 

N.  J. 
TOBIN,  John  F,  (A  1934),  Salesman,  American 

Blower  Corp.,    Rra.  1404,  228  N.   LaSalle  St. 

Chicago,  111. 
TOONDER,  Clarence  L.   (HI  1933),  Air  Cond. 

Engr,,   Sales  Engrg,    Dept.,   Kelvinator   Corp., 

Plymouth  Rd.,  and  (for  mail),  12701  Strattnoor 

Ave.,  Detroit,  Mich. 
TORNQUIST,    Earl    L.    (A    1934),    Supervisor, 

Distribution  Operation  (for  mail),  Public  Service 

Co.  of  North  Illinois,  72  W,  Adams  St.,  Chicago, 

and  465  Parkside  Ave.,  Klmhurst,  111, 
TORR,  Thomas  W.  (M  1933),  Chief  Engr.,  The 

Rudy  Furnace  Co.,  P.  O.   Box  73,  Dowagiac, 

Mich. 
TORRANCE,     Henry     (M     1933),     Pres.,     17,"> 

Christopher  St.,  and  (for  mail),  11:4  East  17th 

St.,  New  York,  N.  Y. 
TOUTON,  R.  IX  (M  1033),  Tech,  Director  (for 

mail),  Bayuk  Cigars,  Inc.,  Ninth  and  Columbia 

Ave,,  Philadelphia,  and  19  Lodges  Lane,  Cymvyd, 

Pa. 
TOWER,    Elwood    S.    (A/    1930),    Engr,,    1114 

Koppers  Bldg.,  and  (for  mail),  1411  WJghtman 

at.,  Pittsburgh,  Pa. 
TRANE,  Reuben  N.*  (A*  1915),  Prea,  (for  mail), 

The  Trane  Co.,  and  126  South  15th  St.,  LaCrosse, 

Wis. 
TRAUGOTT,  Mortimer   (A   1030),   East  Sales 

Mgr.  (for  mall),  Bryant  Heater  &  Mftf.  Co.,  152 

North  15th  St.,  Philadelphia,  and  721  Meeting 

House  Rd.,  Klfcina  Park,  Pa. 
TREADWAY,  QiHmtita  (J  1082),  Sales  Engr.  (for 

mail),  Chrage  Fan  Co.,  707  Security  Bank  Bldg., 

ami  2018  ColHngwood,  Toledo,  Ohio, 
TRIMMER,    Charles    M.    (J    1035:    S    1033), 

Inspection    and    Testing,    Rockland    Light    8s 

Power  Co,,  105  Pike  St.,  and  (for  mall),  28 

Prospect  St.,  Port  Jervfs,  N*  Y, 


TROS&K,  Joseph  J.  (A  mi),  Vlefc-Pres,  and 
Oru  Mgr,  (for  mail),  Vand«*Troske  Co.,  236 
Winter  Ave,  N.W.,  and  233  Brown  St.  S.E., 
Grand  Rapids,  Mich. 

TRUITT,  Joseph  E.  (M  W30;  A  mi),  Pret,» 
Autovent  Fan  &,  Blower  Co.,  1805  N*  ICostner 
Ave.,  Chicago,  III, 

TRULS0N,  Arthur  F.  (M  1830),  Mech,  Engr,, 
1509  W.  Sixth  Si,,  Ashland,  Wla* 

TRUMBOt  SUas  M.  (A  1926),  Sales  (for  mil), 
Buffalo  Forge  Co..  20  N,  Wacker  Dr.,  Chicago, 
and  0^1  Fmnfclia  St.,  Bowttftn  Orove^  III, 

TRUMP,  Charles  G.  (M  1934),  Pm  and  M.  B* 
(for  mail),  Jam«a  Spear  Stove  &  Htg,  Co..  l§33 
Market  St.,  Philacfelphia,  and  608  Baird  M*t 
Marion,  Pa.  » 

TUCELER,  Franfc  N.  (M  l&20}t  FbUd  Ba^r,.  Bf 
Electric  Vtg,  Co.,  Room  UOi  18  Parte  Row, 
Nw  York,  and  (for  mat!)*  ft»  Wlml^r  $t. 
Freepott,  L.  L,  N,  Y. 


ROLL  OF  MEMBERSHIP 


TUCKERMAN,  George  E.  (M  1932),  Mgr, 
Philadelphia  Br.,  Air  Cond.  Div.  (for  mail), 
York  Ice  Machinery  Corp.,  1238  North  44th  St., 
and  6202  Ogontz  Ave.,  Philadelphia,  Pa. 

TURLAND,  Charles  H.  (M  1934;  .4  1930),  Mgr., 
Htg.  4  and  Vtg.  Dept.,  Kipp- Kelly,  Ltd.,  68 
Higgins  Ave.,  and  (for  mail),  325  Centennial  St., 
Winnipeg,  Man.,  Canada. 

TURNAU,  Edmund  H.  (J  1935;  S  1933),  Cadet 
Engr.,  Koppers  Seaboard  By-Product  Coke  Co., 
and  (for  mail),  23  Polifly  Rd.,  Hackensack,  N.  J, 

TURNER,  George  G.  (A  1934),  Western  Repr. 
(for  mail),  Heating  and  Ventilating,  228  N. 
LaSalle  St.,  Chicago,  and  803  Elmwood  Ave., 
Evanston,  111. 

TURNER,  John  (M  1930),  Sales  Engr.  (for  mail), 
Minneapolis-Honeywell  Regulator  Co.,  285 
Columbus  Ave.,  Boston,  Mass.,  and  Contoocook, 
N.  H. 

TURNER,  John  W.  (M  1928),  Chief  Engr.  (for 
mail),  Pacific  Steel  Boiler  Div.,  Box  1488, 
Detroit,  and  26031  Concord  Rd.,  Royal  Oak, 
Mich. 

TURNER,  Mebane  E.  (M  1934),  Mech.  Engr., 
R.  J.  Reynolds  Tobacco  Co.,  and  (for  mail), 
643  Holly  Ave.,  Winston  Salem,  N.  C. 

TURNO,  Walter  G.  W.  (M  1917;  A  1912),  Secy., 
H.  W.  Porter  &  Co.,  Newark,  and  (for  mail), 
71  Lafayette  Ave.,  East  Orange,  N.  J. 

TUSGH,  Walter  (M  1917),  Htg.  and  Vtg.  Engr., 
Tcnney  &  Ohmes,  Inc.,  101  Park  Ave.,  New 
York,  and  (for  mail),  881  Sterling  PL,  Brooklyn, 
N.Y, 

TUTTLB,  George  H.*  (J  1934),  Htg.  Engr.  (for 
mail),  The  Detroit  Edison  Co.,  2000  Second  Ave., 
and  982Q  Belle  Terre,  Detroit,  Mich. 

TUTTLE,  J.  Frank  (M  1913),  Sales  Agent  (for 
mail),  Warren  Webster  Co.,  Kewanee  Boiler 
Corp.,  127  Federal  St.,  Boston,  and  2  Elmwood 
Ave.,  Winchester,  Mass. 

TUVE,  George  L.*  (M  1932),  Asso,  Prof,  of  Mech. 
Engrg,  (for  mail),  Case  School  of  Applied  Science, 
and  1294  Cleveland  Heights  Blvd.,  Cleveland, 
Ohio. 

TWIST,  Charles  F.  (M  1921),  Secy,  (for  mail), 
Ashwell-Twist  Co.,  907  Thomas  St.,  and  2310 
Tenth  Ave.  N.,  Seattle,  Wash. 

TYLER,  Roy  0,  (M  1928),  East  Salea  Mgr,  (for 
mail),  Modine  Mfg.  Co.,  101  Park  Ave.,  New 
York,  and  15  Highbrook  Ave.,  Pclham,  N,  Y. 

TYSON,  William  H.  (M  1928),  Mgr,  of  Engrg. 
(for  mail),  Goodyear  Tyre  &  Rubber  Co.,  Ltd., 
and  "Kipewa"  Codsall  Rcl,  N.R.,  Wolver- 
hampton,  England. 

U 

UHL,  Edwin  J.  (M  1025),  Uhl  Co.,  132  S,  Tenth 

St.,  Minneapolis,  Minn. 
UHL,  WiHiard  F.  (M  1018),  (for  mail),  Uhl  Co., 

132  S,  Tenth  St.,  and  4716  Lyndale  Ave,  S., 

Minneapolis,  Minn. 
XJHLHORN,  W,  J.  (M  1920),  733  S,  Highland 

Ave.,  Oak  Park,  III 
ULLMAN,  Herbert  G.*  (A  1928),  Mgr,  Mech, 

Product  Development  Lab,  .American  Radiator 

Co,,  P,  0.  Box  850,  Second  St.,  Beechwood  Ave,, 

New  Rqchelle.  and  (for  mail),  107  White  Rd., 

Sctrsdale,  N.  Y. 
0RDAHL,  Thomas  H.   (M  1980),  Consulting 

Engr,  (tor  mall).  726  Jackson  PI,  RW,t  and 

150&-44th  St.  N.W.,  Washington,  D,  C. 


VALE,  Henry  A.  L*  (M  1929),  Managing  Director 
(for  mall),  Vale  Co,,  Ltd,,  141-43  Amasffo  St., 
Chriatehureh,  and  241  Ham  Rd.,  Fftttdalton, 
Chrietchurcb,,  New  Zealand. 

VAN  ALBN,  Walter  T.  (AC  1024),  Htg,  and  Sale* 
Ingr.  (for  mail),  1610  Seventh  Are,,  and  1800 
Darlington  Rd,,  Beaver  Falls,  Pa. 


VAN  ALSBURG,   Jerold   H.    (M  1931),  Engr., 

Hart  &  Cooley  Mfg.  Co.,  and  (for  mail),  R.  No.  3, 

Holland,  Mich. 
VANCE,  Louis  G.  (M  1919),  Partner  (for  mail), 

Vance-McCrea  Sales  Co.,  West  27th  and  Sisson 

Sts.,  and  3800  Egerton  Rd.,  Baltimore,  Md. 
VANDERHOOF,  Austin  L.  (A  1933),  (for  mail), 

A.   L.   Vanderhoof,    Inc.,   2341    Carnegie  Ave., 

Cleveland,  and  3120  Yorkshire  Rd.,  Cleveland 

Heights,  Ohio. 
VAN  HORN,  Howard  T.  (A  1933),  Dist.  Mgr., 

Detroit  Stoker  Co.,  1217  McKnight  Bldg.,  and 

(for  mail),  4537  Grand  Ave.,  Minneapolis,  Minn, 
VERMERE,    Earl    J.    (M    1929),    Sales    Engr., 

Kewanee  Boiler  Corp.,  Warren  Webster  &  Co., 

2341  Carnegie  Ave.,  Cleveland,  and  (for  mail), 

2125  Wyandotte  Ave.,  Lakewood,  Ohio. 
VERNIER,  Marcel  G.  (J  1935;  S  1933),  730  Hill 

Ave.,  Willdnsburg,  Pa. 
VERNON,  J.  Rexford   (M  1928;  A   1926),   (for 

mail),  Johnson  Service  Co.,  1355  Washington 

Blvd.,  Chicago,  and  1020  Austin  St.,  Evanston, 

111. 
VETLESEN,  G.  Unger  (M  1930),  3  East  84th  St., 

New  York,  N.  Y. 
VINCENT,  Paul  J.  (M  1931),  Paul  J.  Vincent  Co., 

2133  Maryland  Ave.,  and  (for  mail),  3807  Beech 

Ave.,  Baltimore,  Md. 
VINSON,  Neal  L.  (J  1935;  5  1932),  630  Clyde  St., 

Pittsburgh,  Pa.,  and  (for  mail),  Box  3007,  Lowell, 

Ariz. 
VIVARTTAS,  E.  Arnold  (M  1910),  Consulting 

Engr.,  121  Parkside  Ave.,  Brooklyn,  N.  Y. 
VOGEL,   Andrew    (M   1926),   Engr.    (for  mail), 

General    Electric   Co.,    and    1821    Lenox   Rd., 

Schenectady,  N.  Y. 
VOGELBACH,  Oscar  (U  1923),  23  William  St., 

North  Arlington,  N,  J. 
VOGT,  John  H.  (A  1925),  Mech.  Engr.  (for  mail), 

New  York  State  Dept  of  Labor,  80  Centre  St., 

New  York,  and  87  Grant  Ave.,  Brooklyn,  N.  Y, 
VOGT,  Joseph  B.   (M  1933;  A   1933;  J  1929), 

1304  Grayton  Rd.,  Grosse  Point  Park,  Mich. 
VOISINET,  Walter  E.  (M  1930),  Sales  Repr.  (for 

mail),  Buckeye  Blower  Co,,  250  Delaware  Ave., 

Buffalo,  and  151  Warren  Ave.,  Kenmore,  N.  Y. 
VOLK,  Joseph  H.  (M  1923),  Pres.  and  Treas.  (for 

mail),  Thos.  E.  Hoye  Htg.  Co.,  1906  W,  St.  Paul 

Ave.,  and  2965  South  43rd  St.,  Milwaukee,  Wis, 
VROOME,   Albert  E.    (M  1932),   Engr.,  E.   I, 

duPont    deNemours    &    Co.,    duPont    Bldg., 

Wilmington,  Del.,  and  (for  mail),  412  Morton 

Ave.,  Rutledge,  Pa. 

W 

WACHS,  Louis  J.  (/  1930),  Engr.,  Carrier  Engrg. 
Corp.,  Chrysler  Bldg.,  New  York,  and  (for  mail), 
354  East  21st  St.,  Brooklyn,  N.  Y. 

WAECHTER,  Herman  P.  (A  1930;  J  1927),  Air 
Cond.  Engr.,  York  Ice  Machinery  Corp.,  Brook- 
lyn, and  (for  mail),  89  Sherman  Ave.,  Tompkin&- 
vllle,  N.  Y. 

WAGNER,  A.  M.  (A  1921),  Mgr.  (for  mail), 
American  Radiator  Co.,  1741  W.  St.  Paul  Ave., 
and  1857  N.  Prospect  Ave.,  Milwaukee,  Wia. 

WAGNER,  Frederick  H.,  Jr.  (M  1934),  Mgr.  Air 
Cond.  Dept.,  New  York  Office  (for  mail), 
American  Blower  Corp.,  401  Broadway,  New 
York,  and  1126  Post  Rd.,  Scaradale,  'N.  Y. 

WAITB,  Harry  (A  1929),  1409  North  17th  St., 
Superior,  Wia. 

WAL0ON,  Charles  D.  (A  1632),  Consulting 
Ena;r,,  Spencer  Foundry  Co*,  Penetang,  Ont., 
and  (for  mail),  82  Femdale  Ave.,  Toronto, 
Ont,,  Canada. 

WALSCER,  Alexander  (A  1025),  Br.  Mgr,  (for 
mail),  C.  A,  Dunham  Co.,  Ltd.,  1807  Fifth  SL 
W.f  and  0Q3~13th  Ave.  W.,  Calgary,  Alberta, 
Canada. 

WALKER,  Edmund  R.  (M  1934),  Salea  Mgr., 
Htg.  Div*  (for  mail),  Feddera  Mfg.  Co.,  Inc»  57 
Tonawanda  St.,  and  696  Crescent  Ave.,  Buffalo, 
N.  Y. 


39 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


WALKER,  J.  Herbert*  (M  1916),  Supt.  of 
Central  Htg.  (for  mail),  The  Detroit  Edison  Co., 
2000  Second  Ave.,  Detroit,  and  432  Arlington  Rd., 
Birmingham,  Mich. 

WALKER,  William  Kirby  (M  1935),  Develop- 
ment Engr.,  American  Radiator  Co.,  40  West 
40th  St.,  New  York,  N.  Y. 

WALLACE,  George  J.  (M  1923),  Principal 
96-19-35th  Ave.,  Corona,  and  (for  mail),  27-36 
Ericsson  St.,  East  Elmhurst,  N.  Y. 

WALLACE,  James  Bee  (A  1935),  Dist.  Repr., 
Taco  Heaters,  Inc.,  New  York,  N.  Y.,  and  (for 
mail),  16921  Sorrento  Ave.,  Detroit,  Mich. 

WALLACE,  Kenneth  S.  (M  1931),  Htg.  Engr., 
Peoples  Gas  Light  &  Coke  Co.,  1520  Milwaukee 
Ave.,  and  (for  mail),  5737  Kenmore  Ave., 
Chicago,  111. 

WALLACE,  William  M,  II  (M  1929),  Air  Cond. 
and  Htg.  Contractor,  192  Lexington  Ave.,  New 
York,  and  (for  mail),  S908-19Gth  St.,  Hollis, 
L.  I.,  N.  Y. 

WALLICH,  A,  C.  (M  1919),  Wallich  Ice  Machine 
Co.,  517  E.  Larncd  St.,  and  (for  mail),  1GG7 
Burlingame  Ave.,  Detroit,  Mich. 

WALSH,  James  A.  (A  1932;  J  1929),  Pres.  and 
Mgr.  (for  mail),  James  A.  Walsh,  Inc.,  Architects 
Bldg.,  Philadelphia,  and  Gwynedd  Valley,  Pa. 

WALSH,  J.  Lee  (A  1934),  Sales  Mgr.  and  Engi., 
May  Oil  Burner  Corp.,  Maryland  and  Oliver 
Ave.,  and  (for  mail),  Temple  Ct.  Apts.,  34th  and 
Guilford  Ave.,  Baltimore,  Md. 

WALTERS,  Arthur  L.  (M  1920;  ,4  1925;  J  1924), 
7284  Richmond  PI.,  Maplewood,  Mo. 

WALTERS,  William  T.  (M  1917),  Engr.,  Illinois 
Engrg.  Co.,  Corner  21st  St.  and  Racine  Ave.,  and 
(for  mail),  7905  Phillips  Ave.,  Chicago,  111. 

WALTHER,  Vernon  H.  (M  1928;  J  1925),  Mech. 
Engr,,  6821  Osceola  Ave.,  Edison  Park,  Chicago, 
111. 

WALTERTHUM,  John  J.  (A  1922) ,  Htg.  Con- 
tractor, 173  East  62nd  St.,  New  York,  N.  Y.,  and 
(for  mail),  42-A  Van  Reipen  Ave.,  Jersey  City, 
N.  J. 

WALTON,  Charles  W.,  Jr.  (M  1934),  Mech. 
Engr.  (for  mail),  Rockefeller  Center,  Inc.,  30 
Rockefeller  Plaza,  New  York,  N.  Y,,  and  120 
Monte  Vista  Ave.,  Ridgewood,  N.  J. 

WANDLESS,  Franklin  W.  (M  1925),  Registered 
Engr.  (for  mail),  1518  Fairmount  Ave.,  Phila- 
delphia, and  Berwyn,  Pa. 

WARD,  Frank  James  (M  1935),  The  Frank  J. 
Ward  Co.,  Cold  Spring,  Ky. 

WARD,  Oscar  G.  (M  1919),  Dist.  Repr.  (for  mail), 
Johnson  Service  Co.,  1230  California  vSt.,  and 
1607  Jasmine  St.,  Denver,  Colo, 

WARING,  J.  M.  S,  (M  1932),  Consulting  Kn«r. 
(for  mail),  Chase  &  Waring,  17  East  42nd  St., 
and  277  Park  Ave,,  New  York,  N.  Y. 

WARREN,  Clarence  N.  (M  1919),  Vice-Prea,, 
Hayes  Bros.,  Inc.,  236  W.  Vermont  St.,  and  (for 
mail),  419  East  48th  St.,  Indianapolis,  Ind, 

WARREN,  Francis  C,  (M  1934),  Salesman  (for 

mail),  American  Blower  Corp.,  228  N.  LaSalle 

St.,  Chicago,  and  127  East  Ave.,  Park  Ridge,  111. 

WARREN,  Harry  L.  (M  1930),  130S  Huntington 

Pr,,  South  Pasadena,  Calif. 
WASHBURN,  Marcus  J.  (A  1934),  Insulation 
Engr,  (for  mail).  Eagle- Picher  Lead  Cq.,  Temple 
Bar  Bldg,,  and  2211  Park  Ave.»  Cincinnati,  Ohio, 
WASHINGTON,  G<M>r&e  (M  1934),  Sales 


"for  mail),  Hoffman  Specialty  Co,,  130  N.  Wells 
St.,  Chicago,  and  4327  Johnson  Ave.,  Western 
Springs,  111, 

WASHINGTON,  Laurence  W.  (M  1929),  2301 
Knox  Ave.,  Chicago,  111, 

WATERMAN,  Joh»  H.  (M  1031),  Engr,  (for 
mail),  Chas.  T.  Main,  Inc.,  201  Devonshire  St., 
Boston,  and  7  Centre  St.,  Cambridge,  Masa, 

WATERS,  Georfte  G.  (M  19S1;  A  11)20),  Dist. 
Mgr.  (for  mall),  American  Blower  Corp.,  I4tf8 
Oliver  Bldg,,  and  52  Vernon  Dr.,  Pittsburgh  (l«), 
Pa. 


WATSON,  M.  Barry  (&/  1028),  Consulting  Engr., 

121  Welland  Ave.,  Toronto  5,  Canada. 
WAXJNG,  Tsing  F.  (J  1933),  Htg.  Engr.,  Ander- 
sen, Meyer  &  Co.,  Ltd.,  and  (for  mail),  No.  16, 
Lane  152,  Edinburgh  Rd.,  Shanghai,  China. 
WEBB,   John   S.    (M   1920),    16    Brookline   St., 

Needham,  Mass. 

WEBB,  John  W.  (M  1926),  Managing  Dir.,  Webb 

Dust  Removing  &  Drying  Co.,  Ltd.,  Princess  St. 

Works,  and  (for  mail),  G  Meadows  Rd.,  Heaton 

Chapel,  Stockport,  England. 

WEBSTER,  E.  Kcssler  (M  191,5),  Warren  Webster 

&  Co.,  17th  and  Federal  Sts.,  Camden,  N.  J. 
WEBSTER,    Warren    (Life    Member;    M    1906; 
A    1S99),   Warren   Webster   &    Co.,    17th   and 
Federal  Sts.,  Camden,  N.  J. 

WEBSTER,  Warren,  Jr.  (M  1932;  A  1932; 
J  1927),  Vice-Pres-Treas.  (for  mail),  Warren 
Webster  &  Co.,  17th  and  Federal  Sts.,  Camden, 
and  Washington  Ave.  and  Colonial  Ridge, 
Haddonfield,  N.  J. 

WECHSBERG,  Otto  (M  1932),  Pres.  and  Gen. 
Mgr.,  Coppus  Engrg.  Corp.,  344  Park  Ave.,  and 
(for  mail),  1006  Main  St.,  Worcester,  Mass. 
WEGMANN,  Albert  (A/  1918),  0200  North  17th 

St.,  Philadelphia,  Pa. 

WEIL,  Martin  (A  1925),  Vice-Pres.  (for  mail), 
Weil-McLain  Co.,  641  W.  Lake  St.,  and  4259 
Hazel  Ave.,  Chicago,  111. 

WEIL,  Maurice  I.  (A  1928),  Pres.  (for  mail), 
Chicago  Pump  Co.,  2330  Wolfram  St.,  and  1409 
Elmdale  Ave.,  Chicago,  111. 

WEIMER,  Fred  G.  (A  1919),  Salesman,  Kewanee 
Boiler  Corp.,  1741  W.  St.  Paul  Ave.,  and  (for 
mail),  3958  N.  Stowcll  Ave.,  Milwaukee,  Wis. 
WEINSHANK,    Theodore*    (Life    Member;    A/ 
1906),  (Board  of  Governors,  1913),  3307  Holden 
Ave.,  Chicago,  111. 
WEISS,   Arthur  P,    (A/   192S),   134    Farrington 

Avt\,  North  Tarry  town,  N.  Y. 
WEISS,    Carl    A.    (A    1924),    Supt.    (for    mail), 
Kornbrodt  ICornicft  Ko.,  1811  Troost  Ave.,  and 
29  East  68th  St.,  Kansas  City,  Mo. 
WEITZEL*    Paul    H.    (S    1934),    Cameron    B. 
Weitzel,  and  (for  mail),  122  E.  High  St.,  Man- 
heim,  Pa. 
WELCH,  Louis  A.,  Jr.  (A  1929),  443  Second  St., 

vSchcncctady,  N.  Y. 

WELDY,  Lloyd  0.  (M  1930),  Sales  Kngr.  (for 
mail),  Powers  Regulator  Co.,  2720  Greenview 
Ave.,  and  2846  North  77th  Av«.»  Chicago,  111, 
WELSH,  Harry  S.  (A/  1906) ,  Sales  Engr.,  Weil- 
McLuin  Co.,  and  (for  mail),  53  Kemphurst  Rd., 
Rochester*  N.  Y. 

WELTER,  M.  A.  (A  1925)  (for  mail),  Twin  City 
Furnace  Co.,  410-12  W.  Lake  St.,  and  4«06  £, 
Garfield,  Minneapolis,  Minn, 
WENDT,  Ed&ar  F»  (M  1918),  Prea.  (for  mail), 
Buffalo   Forge   Co.,   400   Broadway,   ancl    120 
Lincoln  Pkwy.,  Buffalo,  N,  Y, 
WEST,   Perry*   (M  1911),   (Council,   19»(M92fi: 
Treas.,  1924-1925),  Consulting  Engr.  (for  mail), 
13  Central  Ave.,  and  445  Ridge  St.,  Newark,  N.  J, 
WETZELL,  Horace  E,  (M  1934),  Chief  Engr,  (for 
mail),  The  Smith  &  Oby  Co.,  0107  Carnegie 
Ave.,  and  8790  KIsraere  Dr.,  Cleveland,  Ohio* 
WHAIXON,   Fletcher  (S  1935),  3852  Lyndale 

Ave.  S.,  Minneapolis*  Minn. 
WHEELEK,  Otto  J.  (M  1923),  Prea-Trets,  (for 
mall),  The  Samuel  A.  Essweln  Htg,  &  Plbg.  Co.. 
548-558  W.  Broad  St.,  and  3044  Collingswood 
Rd.t  Columbus,  Ohio. 

WHELLER,  Harry  S.  (M  1016),  Vice-Pres.,  L,  J. 

Wing  Mfg.  Co*,  154  West  14th  St.,  New  York, 

and  (for  mail),  725  Union  Ave.,  Elizabeth,  N,  J, 

WHITE,  Eu&wfc  B»  (M  1934),    Architect  and 

Engr.  (for  mall],  Y,  M,  C,  A,»  19  S,  USalte  St., 

Chicago,  and  300  N,  Taylor  Ave.,  Oak  Park,  III 

WHITE,   Everett  A.   (M  1021),   Engrg.  Dept., 

Crane  Co.,  30  South  10th  St.,  and  (for  mail), 

5244  Nottingham  St.,  St.  Louis*  Mo. 

WHITE,  El  wood  S.  (M  1921)  t  Pres.  (for  mm*I), 

Taco  Heaters.  Inc.,  Room  1224,  U%  Madiaon 

Ave.,  New  York.  N,  Y.,  and  M«adowbank  Rd,, 

Old  Greenwich,  Conn. 


40 


ROLL  OB  MEMBERSHIP 


WHITE,  John  C.  (M  1932),  State  Power  Plant 
Engr.  (for  mail),  (324  E.  Main  St.,  and  622  E. 
Main  St.  Madison,  Wis. 

WHITELAW,  H.  Leigh  (M  191G),  Vice-Pres.  (for 
mail),  American  Gas  Products  Corp.,  40  West 
40th  St.,  New  York,  N.  V.,  and  Overbrook  Lane, 
Daricn,  Conn. 

WHITELEY,  Stockett  M.  (M  1933),  Consulting 
Engr.  (.for  mail),  Baltimore  Life  Bldg.,  and  3931 
Canterbury  Rd.,  Baltimore,  Md. 

WIIITMER,  Robert  P.  (M  1935),  Secy.,  American 
Foundry  &  Furnace  Co.,  and  (for  mail),  1402  E. 
Washington  St.,  Bloomington,  111. 

WH1TSON,  Lee  S.  (S  1935),  48*1  Harriet  Ave., 
Minneapolis,  Minn. 

WHITTALL,  Ernest  T.  (A  1933),  Vice-Pres.  (for 
mail),  May  Oil  Burner  of  Canada,  Ltd.,  196 
Adelaide  St.  W.,  and  11  Cottingham  Rd., 
Toronto,  Ont.,  Canada. 

WIEGNER,  Henry  B.  (M  1919),  Mgr.,  Boston 
Office,  Johnson  Service  Co.,  20  Winchester  St., 
Boston,  and  (for  mail),  143  Standish  Rd.,  Water- 
town,  Mass. 

WIERENGA,  Peter  O.  (A  1931),  Vice-Pres.  (for 
mail),  C.  C.  James  Co.,  49  Coldbrook  St.  N.E., 
and  231  Brown  St.  S.E.,  Grand  Rapids,  Mich. 

WIGGINS,  Oswald  James  (J  1935;  5  1933), 
Walnut  Grove,  Minn. 

WIGGS,  G.  Lome  (A  1932;  J  1924),  Consulting 
Engr.  (for  mail),  University  Tower,  and  4,797 
Grosvenor  Ave.,  Montreal,  Que.,  Canada. 

WIGLE,  Bruce  M.  (A  1920),  Pres.  (for  mail), 
Bruce  Wigle  Plbg,  &  Htg.  Co.,  9117  Hamilton 
Ave.,  and  18114  Oak  Dr.,  Detroit,  Mich. 

WILDER,  Edward  L.  (M  1915),  Mgr.,  Gas  Sales 
(for  mail),  Utility  Management  Corp.,  120  Wall 
St.  New  York,  and  149  Mt.  Joy  Place,  New 
Roche-lie,  N.  Y. 

WILEY,  Edgar  0.  (M  1909),  Wiley  &  Wilson, 
Lyiichburg,  Va. 

WILIIELM,  Joseph  E.  (S  1934),  1355  West  87th, 
Cleveland,  Ohio. 

WILKINSON,  Farley  J.  (M  1933),  Engr., 
Montgomery  Ward  &  Co.,  Chicago,  and  (for 
mail),  18257  Martin  Ave.,  Homewood,  III. 

WILLARD,  Arthur  Cutts*  (M  1914),  (Presi- 
dential Member},  (Pres.,  1928;  lat  Vice-Pres,, 
1927;  2nd  Vice-Prcs.,  1926;  Council,  1925-1929), 
Prea.  (for  mail),  University  of  Illinois,  President  s 
Office,  and  711  Florida  Ave,,  Urbana,  111. 

WILLEY,  Earl  C.  (M  1934),  Mech.  Kngrg. 
Instructor,  Oregon  State  College,  and  (for  mail), 
1052  "A"  St.,  Curvallia,  Ore, 

WILLIAMS,  Alien  W,  (A  1915),  Managing  Dir- 
ector (for  mail),  National  Warm  Air  Htg,  &  Air 
Conditioning  Assn.,  50  W.  Broad  St.,  Columbus, 
and  f>l  Meadow  Park  Ave.,  Boxley,  Ohio. 

WILLIAMS*  Frank  H.  U  1934),  Air  Cond. 
Tester,  Frigiclaire  Div.,  General  Motors,  Prigi- 
daire  Corp.,  and  (for  mail),  14  Grand  Apts,, 
n,  Ohio, 


WILLIAMS*  J.  McFarland,  Jr.  (A  1928;  J  1927), 
Sales  Kngr,»  1407-3fith  St.  N.WM  Washington, 
D.  C, 


'  WILLIAMS,  J.  Walter  (M  1916),  Prea.  (for  mail), 

Forest  City  Plbg,  Co.,  382-80  E.  State  St.,  and 

928  E.  State  St.,  Ithaca,  N.  Y, 
WILLIAMS,  Leo  E.  (A  1933;  J  1930),  Viscose 

Co,,  and  (for  mail),  827  Liberty  St.,  Meadville, 

Pa, 
WILLIS,  Leonard  L.  (3  1935),  1212  Oliver  N., 

Minneapolis,  Minn. 
WILMOT,  Charles  S,  (M  1919).  (for  mail),  106 

South  16th  St,»  Philadelphia,  ana  406  Essex  Ave.» 

Narberth,  H, 
WXLSON,   Georfi©  T.   (M  1925).   Sales  Engr,, 

Gumey  Foundry  Co..  Ltd.,  4  Junction   Rd., 

Toronto,  and  (for  mail),  Tyre  Ave.,  Islington, 

Qnt.,  Canada, 
WILSON,  Harold  AM  Jr.  (J  1933),  General  Sales 

Dept,  American  Radiator  Co.,  40  W«st  40th  St, 

and  (for  mall),  113B  Park  Am»  New  York,  N,  Y. 


WILSON,  Raymond  W,  (M  1934),  Member  of 
Firm  (for  mail),  Wilson-Brinker  Co.,  412  Pythian 
Bldg.,  and  429  Creston  Ave.,  Kalamazoo,  Mich. 

WILSON,  W.  H.  (A  1932),  Steamfitter  Foreman, 
Pullman  Car  &  Mfg.  Corp.,  11001  Cottage 
Grove  Ave.,  and  (for  mail),  22  West  110th  PL, 
Chicago,  111. 

WILSON,  William  H.  (A  1923),  Br.  Mgr.  (for 
mail),  Johnson  Service  Co.,  507  E.  Michigan  St., 
and  2023  E.  Olive  St.,  Milwaukee,  Wis. 

WINANS,  Glen  IX  (M  1929),  Engr.  of  Steam 
Distribution  (for  mail),  The  Detroit  Edison  Co., 
2000  Second  Ave.,  and  161S3  Wisconsin,  Detroit, 
Mich. 

WINQUIST,  Walter  J.  (A  1930),  Htg.  and  Vtg. 
Engr.,  294  Nostrand  Ave.,  Brooklyn,  N.  Y. 

WINSLOW,  G.-E.  A.*  (M  1932),  Prof,  of  Public 
Health  (for  mail),  Yale  University,  310  Cedar 
St.,  and  314  Prospect  St.,  New  Haven  Conn. 

WINTERBOTTOM,  Ralph  F.  (M  1923),  Htg. 
Engr.,  Winterbottom  Supply  Co.,  Commercial 
and  Miles,  and  (for  mail),  1002  Riehl  St.,  Water- 
loo, Iowa. 

WINTERER,  Frank  C.  (M  1920),  Sales  Mgr.  (for 
mail),  Cochran  Sargent  Co.,  Broadway  and 
Kellogg  Blvd.,  and  830  Juno  St.,  St.  Paul,  Minn. 

WINTHER,  Anker  (J  1932),  Air  Cond.  Engr., 
York  Ice  Machinery  Corp.,  2110  Gilbert  Ave., 
Cincinnati,  Ohio. 

WISE,  Daniel  E.  (S  1934),  10805  Lee  Ave., 
Cleveland,  Ohio. 

WITHER,  Charles  N.  (J  1930),  Dist.  Dealer 
Supv.  (for  mail),  Carrier  Engrg.  Corp.,  2022 
Bryan  St.,  and  4154>£  Prescott  St.,  Dallas, 
Texas. 

WOESE,  Carl  F,  (M  1934),  Consulting  Engr.  (for 
mail),  Robson  &;  Woese,  Inc.,  1001  Burnet  Ave., 
and  256  Robineau  Rd.,  Syracuse,  N.  Y, 

WOHL,  Maurice  W.  (M  1934),  Engr.,  American 
Insulating  Corp.,  377  Atlantic  Ave.,  and  (for 
mail),  32  Lenox  Rd.,  Brooklyn,  N.  Y. 

WOLF,  J.  C.  (M  1923),  Drafting  Room  Engr.  (for 
mail),  B.  F.  Sturtevant  Co.,  and  44  Central  Ave., 
Hyde  Park,  Mass. 

WOOD,  Frederick  G.  (J  1931),  Sales  Air  Cond. 
Kngr.  (for  mail),  Westerlin  &  Campbell  Co. 
(Agents  York  Ice  Machinery  Corp.),  1113 
Cornelia  Ave.,  and  1905  Estes  Ave*,  Chicago,  111. 

WOOD,  J.  Sydney  (M  1926),  Estimator  (for 
mail),  Bennett  &  Wright,  Ltd.,  72  Queen  St.  E., 
and  163  Briar  Hill  Ave.,  Toronto,  Ont.,  Canada. 

WOODMAN,  L.  E,  (M  1934),  Pres.  (for  mail), 
Woodman  Appliance  &  Engrg.  Corp.,  203  E. 
Capitol,  and  1014  Fairmount,  Jefferson  City, 
Mo. 

WOODRUFF,  Wilbur  3,  (M  1933),  Woodrutf 
Coal  Co,,  206  N.  Broadway.  Urbana,  111. 

WOODS,  Edward  H.  (M  1934),  Prop,  (for  mail), 
F.  H.  Higgins,  311  E.  State  St.,  and  Hook  PL, 
Ithaca,  N.  Y. 

WOOLLARD,  Mason  S.  (M  1934),  Draftsman. 
H.  H.  Angus  Consulting  Engr.,  1221  Bay  St.,  and 
(for  mail),  31  Hillcrcat  Park  Ave,,  Toronto,  Ont., 
Canada, 

WOOLSTON,  A,  H.  (M  1919),  2015  Sansom  St., 
Philadelphia,  Pa. 

WORSHAM,  Herman  (M  192d;  J  1018),  Frial- 
daire  Sales  Corp.,  Dayton,  Ohio,  and  (for  man), 
103  N,  Walnut  St.,  East  Orange,  N.  J, 

WRIGHT,  Clarence  E.  (J  1935;  5  1033),  Part- 
time  Instructor  in  Htg.,  Carnegie  Institute  of 
Technology,  Carnegie  Tech.,  and  (for  mail),  270 
N,  Bellefield  Ave*,  Pittsburgh,  Pa. 

WRIGHT,  Kenneth  A.  (M  1921),  Johnson 
Service  Co,(  1113  Race  St.,  Cincinnati,  Ohio., 
and  113  Orchard  St.,  Ft.  Mitchell,  Ky.  , 

WRIGHT,  M,  Bfcrney  (A  1982;  J  1929),  Astt. 
Prof,  of  Mech.  Enjfrg.  (for  mail),  The  Drexel 
Institute,  Philadelphia,  and  228  Essex  Ave., 
Narberth,  Pa, 

WUNDERLICH,  Milton  $»*  (M  1925),  I&iulit* 
Co.,  Minneapolis,  and  (for  mail),  1^8  Laurel 
Ave,,  St.  Paul,  Minn, 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


WYLIE,  Howard  M.  (M  1925;  J  1917),  Vice- 
Pres.  in  charge  of  Sales  (for  mail),  Nash  Engrg. 
Co.,  and  51  Elrnwood  Ave,,  South  Norwalk, 

WYMORE,  Fred  C.  (J  1935;  S  1933),  100  West- 
port  Rd.,  Kansas  City,  Mo. 


YAGER,    John   J.    (M   1921),   425  Woodbridge 

Ave.,  Buffalo,  N.  Y. 
YAGLOU,    Constantin    P.*    (M    1923),    Asst. 

Prof,  of  Industrial  Hygiene  (for  mail),  Harvard 

School  of  Public  Health,  55  Shattuck  St.,  Boston, 

and  10  Vernon  Rd.,  Belmont,  Mass. 
YARDLEY,  Ralph  W.   (U  1920),  Asst.  Archt., 

Board  of  Education,  City  of  Chicago,  228  N. 

LaSalle  St.,  Room  568,  Chicago,  111.,  and  (for 

mail),  c/o  Judge  J.  W.  Galbraith,  Farmers  Bank 

Bldg.,  vSuite  601,  Mansfield,  Ohio. 
YATES,  Geor&e  L.  (S  1934),  Oklahoma  University, 

Albert  Pike  Hall,  Norman,  and  (for  mail),  1220 

Johnstone,  Bartlesville,  Okla. 
YATES,  James  E.   (M  1934),   Mgr.   (for  mail), 

Yates,  Neal&Co.,  231  Tenth  St.,  and  431-16th 

St.,  Biandon,  Manitoba,  Canada. 


YATES,  Walter  (Life  Member;  M  1902),  Govern- 
ing Dir.  (for  mail),  Matthews  &  Yates,  Ltd., 
Cyclone  Works,  and  Parksend  Swinton,  Man., 
England. 


ZACK,  Hans  J.  (M  1928),  Zack  Co.,  2311  Van 
Buren  St.,  Chicago,  111. 

ZIBOLD,  Carl  Edward  (M  1929),  Mech.  Engr., 
Htg.  and  Vtg.  Co.,  Colonial  Terrace,  West- 
minster Ridge,  White  Plains,  N.  Y. 

ZIESSE,  Karl  L.  (A  1931),  Secy-Treas.  (for  mail). 
Phoenix  Sprinkler  &  Htg.  Co.,  115  Campau 
Ave.  N.W.,  and  315  Hampton  Ave.  S.E.,  Grand 
Rapids,  Mich. 

ZIMMERMAN,  Alexander  H.  (A  1930),  Venti- 
lation Engr.,  Chicago  Board  of  Health,  707 
City  Hall,  and  (for  mail),  5147  N.  St.  Louis 
Ave,,  Chicago,  111. 

ZINK,  David  D.  (M  1931),  Consulting  Engr.  (for 
mail),  Zink  Home  and  Bldg.  Service,  225  Pkwa 
Theatre  Bldg.,  Kansas  City,  and  Hickman's 
Mill,  Mo. 

ZOKELT,  C.  G.  (M  1921),  ConsultiriK  Engr., 
381Q-24th  Ave.  S.,  Seattle,  Wash. 

ZUHLKE,  William  R.  (M  1928),  530  McLean 
Ave.,  Yonkcrs,  N.  Y. 


Summary  of  Membership 

(Corrected  to  January  1,  1935) 


UNITED  STATES 


Alabama 1 

Arizona 3 

Arkansas 3 

California 34 

Colorado „  5 

Connecticut 30 

Delaware 5 

District  of  Columbia 16 

Florida 3 

Georgia 6 

Illinois. 195 

Indiana 18 

Iowa 6 

Kansas 2 

Kentucky.... 6 

Louisiana 3 

Maine -  4 

Maryland 21 

Massachusetts 103 

Michigan . 94 

Minnesota 117 

Missouri 81 


Montana 3 

Nebraska... 2 

New  Jersey 100 

New  York 357 

North  Carolina 7 

North  Dakota 1 

Ohio 101 

Oklahoma 27 

Oregon 1 

Pennsylvania 210 

Rhode  Island 5 

South  Carolina 2 

Tennessee 6 

Texas 21 

Utah 1 

Vermont 3 

Virginia 12 

Washington 24 

West  Virginia 5 

Wisconsin , 50 


1694 


FOREIGN  COUNTRIES 


Australia 3 

Belgium.... 1 

Canada..... „ 89 

China 

Czechoslovakia  „ fw, 1 

Denmark. , *»..,.....„„.  1 

England .  16 


France.. 

Germany. 
Holland. 
India...... 

Ireland 

Italy 


Japan *.,„ 5 

Mexico. 3 

New  Zealand 3 

9     Norway . 2 

Phillipine  Islands 1 

Scotland , 1 

South  Africa ,..„.*.  1 

Sweden » 3 

n  s.  a  PL,.-™... i 


Total  Membership. — .... 


156 


,1850 


SUMMARY  OF  MEMBERSHIP  BY  GRADES 

Honorary  Members ...«„« .»,«.       2 

Presidential  Members ...... 23 

Members,..,.... «,«»,.........«»....,....... .« « *.  1156 

Associate  Members... , 539 

Junior  Member**.... .,«. 233 

Student  Members*..,*,, , * 97 


1850 


LIST  OF  MEMBERS 
Geographically  Arranged 

UNITED  STATES 


ALABAMA 

San  Francisco  — 

Stamford  — 

ILLINOIS 

•n                  A      T 

Wnirf     T      W 

Birmingham  — 

Cochran,  L.  H. 
Corrao,  J. 

Torrington  —                    Blooming  ton  — 

Lichty,  C.  P. 

Haley,  H.  S. 

Doster,  A.                           MaGirl,  W.  J. 

Hudson,  R.  A. 

soper,  H.  A. 

ARIZONA 

Krueger,  J.  I. 

Waterbury  —                         Whitmer,  R.  P, 

Leland,  W.  E, 

Scott,  W.  P.,  Jr. 

Ahlberg,  H.  B.                 Qh 

icago  — 

Tucson  — 

Moreau,  D. 
Phoenix- 

South  Pasadena  — 

Warren,  H.  L. 

Hutzel,  H.F.                      Acoeriy,  j.  j. 
Simpson,  W.  K.                 Arenberg,  M.  K. 
Stewart,  C.  W.                    Baurngardner,  C.  M 
Black,  F.  C, 

Keys,  L.  F. 

COLORADO 

Holte,  E'  E  ' 

T 

DELAWARE                  Borlinir.  t7p 

Lowell*-*- 

Bracken,  J.  H, 

Vinson,  N.  L. 

Colorado  Springs  •* 

• 

t^aun.  L.  1\ 

ARKANSAS 

Davis,  A.  F. 
Jardine,  D.  C. 

Wilmington—                      Broom,  H.  A. 

Fort  Smith— 

Herrick,  L. 
Pine  Bluff— 

Denver:  — 

McQuaid,  D.  J, 
O'Rear,  L.  R. 
Ward,  0.  G. 

Lownsberry,  B.  F,               rhS'JJuH^  T' 
Schoenijalm,  R.  P.              $S?p"b.W<  K 
Cron<?.  (1  Tt.    TT- 

Greer,  W.  R. 

CONNECTICUT 

DISTRICT  OF               < 

Cunningham,  T.  M. 

Siloam  Springs  — 



COLUMBIA 

Cutler,  J,  A. 

Jones,  C.  R, 

Bridgeport— 

•-•'V.UU,     i1  ,    J  ft      1  I', 

—  -  —  -                      DeLand,  C.  W. 

CALIFORNIA 

Faile,  K,  H, 
Smak,  J.  R, 

¥,y    ,  ,                                   Doherty,  R. 
Washington—                     Dunham,  C,  A, 

Beitzell,  A.  E.                    3 

iramert,  L.  1>, 

Fairfield— 

Brown,  W»  A, 

Cric8«<m,  K.  B. 

Berkeley  — 

Osborn,  W,  J. 

Coward,  H.                        Kvleth,  K.  B. 

Duncan,  G,  W.,  Ji. 
Beverly  Hills- 

Greenwich  — 

Jones,  A.  L. 

Downea,  H.  II.                   Finan,  J,  J, 
Febrey,  E.  J.                      KiUaeruld,  M.  J. 
Feltwell,  R.  H.                   Fleming,  J,  P. 

Nelson,  H.  A. 

Oppcrman,  E.  F. 

Frankel,  G.  S.                    Foster,  T.  R, 

Huntin&ton  Park— 

Manchester— 

Gardner,  S,  F.                    Frank,  J,  M, 
Hood,  0.  P.                       Gardner,  W..  Jr> 

Barman,  W,  E,,  Jr. 

Buck,  L, 

Kiczalea,  M.  0,                  Gaylord,  F,  H. 

Los  Angeles— 
Anderson,  C.*S, 

Millard,  J.  W, 
New  Britain— 

Mayette,  C.  E,                   Getschow,  G.  M, 
Ourusoff,  L.  S.                   Getschow,  R.  M, 

Smallman,  E.  W.               Gibbs,  F,  C, 

Binford,  W,  M. 
Bullock,  H,  H, 

Hjerpe,  C.  A.,  Jr. 

Thompson*  N.  S.               Goelz,  A,  IL 
Urdahl,  T,  H»                    Goaaetl.  E.  .1, 

Cranston,  W,  E,,  Jr. 
EUingwood,  E.  L. 
Hill,  F.  M, 
Holliday,  W.  L, 
Hungerford,  L. 
Kendall,  E,  H, 
Kennedy,  M. 

New  Haven— 
Greenburg,  L. 
Hughes,  C,  K. 
Seeley,  L,  E. 
Teasdale,  L,  A. 
Winslow,  C.-E.  A. 

Williams,  J.  M,,  Jr, 
FLORIDA 

Graves,  W,  B. 
laag,  S.  L. 
iainesi,  J.  J, 
tiale,  J.  R 
Flanley,  T,  F,t  Jr. 
Iwt.H,  M, 
Htattii,  R,  K, 

,._„,,   , 
Ft.  Lauderdale— 

Kooistra,  J.  F. 
Ness,  W,  H.  C. 
Drear,  A,  G, 
Ott,  0.  W. 
Park,  J,  F, 

New  London— 
Chapin,  C.  G, 
Forsberg,  W, 
Hopson,  W.  T. 

Charlton,  J,  F*                   : 
West  Palm  Beach— 

Htayden,  C.  F, 

g»mJ;J.   B 

Wayward.  R,  B, 
Hteckd,  E,  P. 

f-fjarUWs*    t     t 

Pierce,  E.  D, 

Norwaik— 

pto!Tbw'L"         HiTCEv:" 

Polderman,  JU  H, 
Scofield,  P,  C, 

Mead,  E,  A, 

Hlncldey,  H.  B, 

Simonds,  A.  H, 
Oakland— 

Riverside-- 
Murphy, J.  R» 

Horton,  1C  F, 
—                     Howttt,  J, 
.   .      ..                                   HowelL  L. 

Cummings,  G,  J, 
Pasadena— 

South  Norwalfc— 

Adams,  H.  E, 

Atlanta—                         ] 

Clare,  F.  W, 
Kent,  L,  F, 

Hiubbard,  G,  W. 
rluttoel,  A,  U. 
enninp,  W,  G. 

Giftord,  R,  L, 

Harvey,  A*  D. 

Klein,  E,  W. 

toion,  L  S. 

San  Diego 

Sadler,  C.  B, 

Jennings,  L  C. 
Lyona,  C,  J, 
Wylte,  H,  M, 

McKinney,  W,  J, 
T«mplln,  C.  L, 
Thornton,  W,  B,               ! 

'ohns,  H.  B, 
'ohnton,  C»  W, 

£*****  A*»    w    T& 

vcency,  j?  .  t  . 

44 


ROLL  OF 

MEMBERSHIP 

Kehm,  H.  S. 

Wood,  F.  C. 

INDIANA 

LOUISIANA 

Kreiasl,  H.  G. 

Zack,  H.  J. 

Lagodzinski,  H.  J. 
Lautenschlager,  F, 

Zimmerman,  A.  H. 

Evansville  — 

Choudrant  — 

Lawler,  M.  M. 

Cicero  — 

Bulleit.  C.  R. 

Rinehart,  W.  R. 

Lees,  H.  K. 
Lewis,  S.  R. 

Keppner,  H.  W, 

Huntington  — 

New  Orleans  — 

Lockhart,  H.  A. 
Machen,  J.  T. 

Decatur  — 

Smith,  G.  W. 

Gammill,  0.  E.,  Jr. 
May,  G.  E. 

Malone,  D.  G. 

Shorb,  W.  A. 

Indianapolis  — 

Malvin,  R.  C. 
Marschall,  P.  J. 
Martin,  A.  B. 
Matchett,  J.  C. 
Mathis,  E. 

Elmhurst— 

Potvin,  L.  J. 

Evanston  — 

Ammerman,  C.  R. 
Fenstermaker,  S.  E. 
Hagedon,  C.  H. 
Hayes,  J.  G. 
Kruse,  R.  W. 

MAINE 

Auburn  — 

Fowles,  H.  H. 

Mathis,  H. 
Mathis,  J.  W. 

Hayes,  J.  J. 
Moore,  R.  E. 

Ott,  R.  C. 
Poehner,  R.  E. 

Portland  — 

Mathis,  V.  J. 

Shivers,  P.  F. 

Pels,  A.  B. 

Mauer,  W.  J. 

Homewood  — 

Supple,  G.  B. 

Merrill,  C.  J. 

May,  M.  F. 
McCauley,  J.  H. 

Wilkinson,  F.  J. 

Warren,  C.  N. 

Swanson,  H. 

McClellan,  J.  E. 
McDonnell,  E.  N. 
McFarland,  W.  P. 

Joliet  — 

Russell,  W.  B. 

Lafayette  — 

Hoffman,  J.  D. 

MARYLAND 

Mcllvaine,  J.  H. 
Mert/,  W.  A. 

Kewanee  — 

LaPorte  — 

Baltimore  — 

Miller,  F,  A. 

Bronson,  C.  E. 

Shrock,  J.  H. 

Axeman,  J.  E. 

Miller,  J.  E. 
Miller,  R.  T. 
Milliken,  J.  H. 

Dickson,  R.  B. 
HaitmanXJMM. 
Pursell,  H.  E. 

Michigan  City— 

Stockwell.iW.  R. 

Collier,  W,  I. 
Hall,  M.  S. 
Lednum,  J.  M, 

Mittendorff,  E.  M, 
Mueller,  H,  C. 
Murphy,  E.  T. 
Narowetz,  L.  L.,  Jr. 

LaGrange  — 

Eaton,  B.  K. 

Peru— 

Pyle,  J.  W. 
Thrush,  H.  A. 

Leilich,  R.  L. 
Maccubbin,  H.  A. 
McCormack,  D. 
Morris,  E.  J. 

Needier,  J.  H, 
Neiler,  S,  G. 

Linn,  H.  R. 

St.  Mary-of-the- 

Posey,  J. 
Seiter,  J.  E. 

Nelson,  C.  L, 

Moline— 

Woods 

Shepard,  J,  deB. 

Newport,  C.  F. 
O'Brien,  J.  H. 

Beling,  E.  H. 
Nelson,  H.  W. 

Bisch,  B.  J. 

Smoot,  T.  H. 
Vance,  L.  G. 

Often,  B. 

Nelson,  R.  H. 

IOWA 

Vincent,  P.  J, 

Olscn,  C.  F. 

Otis,  G.  K. 

Walsh,  J.  L. 

Peller,  L. 

Whiteley,  S.  M. 

Pitcher,  L.  J  . 
Pope,  S.  A. 
Powers,  F.  W. 

Mt.  Vernon— 
Benoist,  L.  Lt 

Ackley— 

Nelson,  G.  O. 

Brooklyn  Park  — 

Rodgers,  J.  S. 

Prentice,  (),  J. 
Price,  C,  K. 
Rusmuasen,  R.  P. 

Norwood  Park  — 

Olson,  B. 

Cedar  Rapids- 
Chandler,  C.  W. 

Chevy  Chase  — 

Dalla  Valle,  J.  M, 

Raymond,  F.  I. 
Reger,  H.  P, 
Reid,  H.  P. 

Oak  Park— 

Blanding,  G,  H. 

Council  Bluffs  — 

Huffacker,  H.  B. 

Ellicott  City— 

Tibbets,  J.  C. 

Ries,  L.  S. 

May,  E,  M. 

LeMars  — 

Rockville  

Rieu,  E,  W. 
Rottmayer,  S,  L 

Nightingale,  G.  F. 
Uhlhorn,  W.  J. 

Mathey,  N.  J. 

Brunett,  A,  L, 

Sawhill,  R,  V. 
Scheideeker,  D.  B, 
Schweim,  H.  J. 

Peoria— 

Farnsworth,  J.  G, 

Sioux  City— 

Hagan,  U.  V. 

Roland  Park— 

Dorsey,  F.  C. 

Seelig,  L, 
Shultz,  E. 
Small,  J.  D. 

Meyer,  F.  L, 
Robb,  J.  M, 

Waterloo— 

Winterbottom,  R.  F. 

MASSACHUSETTS 

Snider*  L.  A. 
Spielman,  G,  P. 

Rocfcford— 

KANSAS 

Arlington  — 

Stannard,  J.  M, 

Braatz,  C,  J. 

Shaw,  N.  J.  H, 

Sutcliffe,  A,  G, 
Thhm,  C,  A, 

Dewey,  R.  P, 

Merwin,  G.  E. 

Hutchinson  — 

Arlington  Heights— 

Thomas,  L,  G,  L. 

Sharp,  H.  C, 

Stevens,  H.  L, 

Tarr,  H.  M. 

Thomas,  R.  H, 
Thommen,  A.  A. 
Tobin,  J,  F. 

Stewart,  D»  J. 
Urbana  — 

Salina— 

Ryan,  W,  F. 

Boston- 
Archer,  D.  M. 

Tornqulst,  E.  L, 
Truitt,  J,  E. 

Brodericfc,  E.  L. 
Fahnestock,  M.  K. 

KENTUCKY 

Bartlett,  A.  C, 
Berchtold,  E,  W, 

Trumbo,  S.  M. 

Konzo,  S. 

Boyden,  D.  S. 

Turner,  G.  G. 
Vernon,  J.  R, 

Wallace,  K,  S* 

KraU,  A.  P. 
Severna,  W.  H, 
Willard,  A.  C. 

Cold  Spring—           ' 
Ward,  F.  J. 

Brinton,  J.  W. 
Brissettet  L.  A, 
Bryant,  A,  G. 

Walters,  W,  T, 
Walther,  V.  H. 
Warren,  F.  C, 

Woodruff,  W,  J, 
Villa  Park— 

Ft,  Thomas  — 

Stevens,  W.  R. 

Bullock,  T,  A. 
Cumminga,  C.  H. 
Drinker,  P. 

Washington,  G. 
Washington,  L,  W. 

Armapach,  0,  W, 

Lexington  — 

Dusossoit,  E.  A. 
Edwards,  D.  J. 

Well,  M, 
Weil,  M,  L 
Weinihank,  T. 
Weidy,  L.  0, 
White,  E.  B, 

Wllinette— 

Norris,  W.  0, 

Wlanetka— 

O'Bannon,  L.  S. 
Louisville  — 

Helbtrom,  J. 
Murphy,  H.  C, 

Foulds.  P,  A.  L. 
Franklin,  R.  S. 
Gleason,  G.  H. 
Hajek,  W,  J* 
Herrick*  D,  A. 

Wilson,  W.  H, 

Ellis,  B,  B, 

Reed,  W,  M. 

Hilliard,  C,  E, 

AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

Hoyt,  C.  W. 

Melrose  — 

Detroit— 

Holland— 

Keefe,  E.  T. 

Cole,  E.  Q, 

Akers,  G.  W. 

Cherven,  V.  W. 

Kelley,  J.  J. 

Dodge,  H.  G. 

Arnoldy,  W.  F. 

Sawyer,  J.  N. 

Kellogg,  A. 

Gerrish,  G.  B. 

Baldwin,  W.  H. 

Van  Alsburg,  J.  H. 

Kimball,  C.  W. 

Barth,  H.  E. 

McCoy,  T.  F. 

Milton— 

Bishop,  F.  R. 

Houfihton  — 

Merrill,  F.  A. 
Miller,  J,  F.  G. 
Moulton,  D. 

Mitchell,  C,  H. 
Needham  — 

Blackmore,  F.  H. 
Boalea,  W.  G. 
Booth,  H.  N. 

Seeber,  R.  R. 
Kalamazoo  —  • 

Osborne,  M.  M. 

Park,  C.  D. 

Brennan,  J.  W, 

Brinker,  H.  A. 

Plunkett,  J.  H. 
Rydell,  C.  A. 

Webb,  J.  S. 

Chappell,  H.  D. 
Chester,  T, 

Downs,  S.  H. 
McConner,  C.  R. 

Shaw,  E. 

Newton  Center  — 

Collamore,  R. 

Temple,  W.  T. 

Stetson,  L.  R. 
Swaney,  C.  R. 

Murray,  J.  J. 

Connell,  R.  F. 
Coon,  T.  E. 

Wilson,  R.  W, 

Turner,  J, 
Tuttle,  J.  F. 
Waterman,  J.  H. 
Yaglou,  C.  P. 

Newton  ville  — 

Emerson,  R.  R. 
McMurrer,  L.  J. 

Cooper,  F.  D. 
Cummins,  G,  H, 
Darlington,  A.  P. 
Dauch,  E,  O. 

Adams,  E.  1. 
Distel,  F. 
McLouth,  B.  F. 

Brookline  — 

Riley,  E.  C. 

Quincy  — 

Gesnier,  J. 
Stone,  E.  R. 

Dubry,  E.  E. 
Egglcston,  L.  W. 
Feely,  F.  J. 
Giguerc,  G.  H. 

Parsons,  R.  A. 
Muskegoii  Heights  — 

Reid,  H.  F. 

Cambridge  — 

Baker,  R.  H, 
Flint  C  T 

Reading  — 

Ingalls,  F,  D.  B. 

Glanz,  E. 
Hamlin,  H.  A. 
Hare,  W.  A. 

St.  Joseph  — 

Milliken,  V.  D. 

Haddock,  L  T, 

Revere  — 

Harms,  W.  T. 

Holt,  J.  W. 
Ivahan,  C. 

Foulds,  S.  T,  N. 

Harrtgan,  E.  M. 
Heydon,  C.  G. 

MINNESOTA 

Lees,  J.  T. 
MacDonald,  E.  A. 

Cochltuate  — 

Ahearn,  W.  J. 

South  Hadley— 

Colby,  C.  W. 

South  wick—- 
Shaw, B.  E. 

Kilner,'  J/S/ 
Kirkpatrick,  A,  H, 
Knibb.  A.  E, 
Luty,  D.  J. 
Maler,  G.  M. 

Cloquot  — 
SpafTord,  A, 
Duluth— 

McColl,  J,  R. 

Foster,  C.. 

Dal  ton  — 

Da  kin,  H.  W. 

Springfield-™ 

Brown,  W.  M. 

McConachie,  L.  L. 
McIntSre,  j.  F. 

Minneapolis— 

Aikcn  J   F 

E.  Dedham  — 

Higgins,  T,  J. 

Cross,  R.  E. 
Holmes,  R,  E. 
Leland,  W.  B. 

McLean,  D. 
Milward,  R.  K, 
Morse,  C.  T, 

Algrcn,  A,  B. 
Anderson,  XX  B. 

Dorchester- 

Murphy,  W.  W. 
O'Neil,  J.  M. 

Olson,  R.  G. 
Paetz,  H.  E. 

Armstrong,  R.  W. 

Brown,  M. 
Goodrich,  C.  F. 
Hosterman,  C.  O. 
Shaer,  I,  E. 

Waban— 

Jones,  W.  T. 
Watertown  — 

Parrott,  L.  G, 
Partlan,  J.  W, 
Purcell,  F.  C. 
Purcell,  R.  E. 
Randall,  W,  C. 

Bctta,  H.  M. 
Bjerken,  M,  H. 
Bredeaen,  B,  P. 
Bull,  A.  S. 

Fltchburg— 

Wiegner,  H.  B. 

Rowe,  W.  A. 
Sanford,  S,  S, 

BuotF  A*  V. 
Burns,  JtC.  J. 

Karlson,  A.  F. 
McKittrick,  P,  A. 

Wollesley  Hills- 
Barnes,  W.  E. 

Sauer,  R.  L. 
SchHchting,  W.  G. 

Burritt,  C.  G. 
Butts,  R,  L. 

Harwich  Port  — 

Gilling,  W.  F.,  Jr, 

Shea,  M,  B. 

Copperud,  E,  R. 

Maxwell,  G,  W. 
Holbrook— 

Nason,  G.  L, 

Hyde  Park- 
Ellis,  F.  R. 
Epple,  A,  R. 
Fritzberg,  L,  H, 
Keyes,  R.  K, 
Wolf,  J,  C, 

Ipswich--*- 

Monroe,  R.  R. 

West  Roxbury—- 

Christie,  A,  Y, 
McPherson,  W,  A. 

Weymouth  — 
Clough,  L. 
Winchester—- 
Jackson, A,  B. 
Woburn— 
Parker,  P. 
Wollaaton— 
Hodgdon,  H.  A, 

Snyder/J.  W, 
Spifcsley,  R,  L. 
Spurgeon.  J.  H. 
Taylor,  W,  E. 
Taze,  0,  L. 
Toonder,  C.  L, 
Turner,  J.  W. 
Tuttle,  G.  H. 
Walker,  J.  H, 
Wallace,  J,  B. 
WalHch,  A,  C. 
Wigle,  B,  M, 
Winant,  G.  D, 

DahMrom,  G.  A, 
Dovolis,  N,  J, 
Forfar,  D,  M, 
Gelb,  A. 
Gcrrigh,  H.  E. 
Gordon,  E.  B.,  Jr. 
Gordon,  W,  J»»  Jr. 
Gross,  L.  C, 
Haatvedt,  S.  R. 
Hall,  J,  R. 
Hanson,  L,  P, 
Han-la,  J.  B, 
Hildebrandt,  H.  A. 
Hitchcock.  P»  C. 
Howard,  E. 

Bride,  W,  T* 

Worcester— 

Robinson,  H.  C. 

Firestone,  J,  F, 
Torr,  T.  W. 

Johnson,  L,  H. 
Jonea,  N,  W. 

JLeomtoter  — 

Wechaberg,  0. 

E.  Lart«I«t&— 

Jordan,  L»  E» 
Jordan,  R.  C. 

Kern,  R,  T. 

Miller,  L»  G, 

KJtngt  R,  L, 

MICHIGAN 

Grand  Rapid*-— 

Bmdfteld,  W,  W, 
Leigh,  R.  L, 

Kntttdtaoru  C.  M» 
Kuehn,  W.  C, 
KuempeL  L,  L, 
Ktins,  Jt  W. 

Brigham,  F.  H. 

Ann  Atbor  — 

Lvntt-— 

Backus,  T,  H,  L, 

Morton,  C.  H. 

Fselmn,  J,  B, 

Battle  Creek— 

Troske,  J,  J. 
Wierenga,  P,  0, 

Leslie,  !0,  B» 

Gates,  W,  A. 

Christenson,  H. 

Ziesae,  K.  L. 

Lewis,  C.  B, 

Medford— 

! 

Grosso  Point  Park*"" 

Lowe,  H,  ». 

Lundi  C»  K» 

Citshman,  L,  D, 

ftadjltfcy,  J.  N, 

Vogt,  J,  B. 

ROLL  OF 

MEMBERSHIP 

Magney,  G.  R. 

MISSOURI 

Rosebrough,  R.  M. 

Elizabeth- 

Martenis,  J.  V. 
Maynard,  H.  R. 

Sodemann,  P.  W. 
Sodemann,  W.  C.  B. 

Burke,  J.  J. 
Cherne,  R.  E. 

Miller,  H.  A. 

Ferguson  — 

Stammer,  E.  L. 

Cornwall,  G.  L 

Miller,  L.  B. 
Mjolsnes,  L.  0. 

Szombathy,  L.  R. 

Tenkonohy,  R.  J. 
White,  E.  A. 

Grant,  W.  A. 

Morgan,  G.  C. 
Morton,  H.  S. 
Myers,  C.  R. 

Jefferson  City  — 

Woodman,  L.  E. 

Webster  Groves  — 

Myers,  G.  W.  F. 

Merle,  A. 
Wheller,  H.  S. 

Nelson,  R.  A. 

Kansas  City  — 

Ranck,  G.  L. 

Essex  Fells— 

Noble,  T.  G. 
Oatrin,  A. 
Pappenfus,  W.  G. 
Pfeifer,  O.  J.,  Jr. 
Porter,  H,  M. 

Adams,  C.  W. 
Allan,  N.  J. 
Arthur,  J.  M. 
Betz,  H.  D. 

Wellston— 

Seepe,  P.  E. 

Stacey,  A.  E.,  Jr. 

Grantwood  — 

Butler,  P.  D. 

Powell,  K.  A. 
Priester,  G.  B. 

Bliss,  G.  L. 
Buckley,  M.  B. 

MONTANA 

Hackensack  — 

Pung,^  D.  W. 
Roberts,  J.  R. 
Rossi  ter,  P.  A. 

Caleb,  D. 
Campbell,  E.  K. 
,     Campbell,  E.  K.,  Jr. 

Big  Timber  — 

Strickland,  A.  W. 

Turnau,  E.  H. 
Haddonfield— 

Rowley,  F.  B. 

Chase,  L.  R. 

Dobbs,  C.  E. 

Sanford,  A.  L. 
Schernbeck,  F.  H. 
Steinkellner,  E.  J. 

Clegg,  C. 
Cook,  B.  F. 
Dawson,  T.  L. 

Billings— 

Cohagen,  C.  C. 

Jones,  R.  E. 
Moody,  L.  E. 

Stiller,  F.  W. 
Sunclcll,  S.  S. 
Sutherland,  D.  L. 

Disney,  M.  A. 
Dodds,  F.  F. 
Downes,  N.  W. 

Bozeman  — 

Powers,  F.  I. 

Hasbro  uck  Heights  — 

Goodwin,  S.  L. 

Swanson,  R.  G. 

Eppright,  J,  O. 

Hawthorne  — 

Sweivcn,  C.  E. 
Swenson,  J.  E. 

Fehlig,  J.  B. 
Filkins,  H.  L. 

NEBRASKA 

Lawton,  F.  C. 

Tabor,  C.  B. 
Uhl,  K.  J. 

Flarsheim,  C.  A. 
Gillham,  W.  E. 

Clarks  — 

Irvington  — 

Uhl',  W.  F, 
Van  Horn,  H.  T. 

Haas,  E.,  Jr. 
Kell,  W.  R, 

Manning,  W.  M. 

Freas,  R.  B. 
Reinke,  A,  G. 

Welter,  M.  A. 
Whullon,  F. 
Whitson,  L.  S. 
Willis,  L,  JL. 

Kitchen,  J.  H. 
Lewis,  J.  G. 
Maillard,  A.  L. 
Matthews,  J.  E. 

Scotts  Bluff— 

Davis,  O.  E. 

Jersey  City  — 

Hasagen,  J.  B. 
Jones,  H.  L. 

Owatonna— 

Millis,  L.  W, 
Natkin,  B. 

NEW  JERSEY 

Kelly,  C.  J. 
Ritchie,  W, 

Clarkaon,  W.  B. 

Nottberg,  G. 
Nottberg,  H. 

Arlington  — 

Schwartz,  J. 
Walterthum,  J.  J. 

Robbimdale— 

Qlchoff,  M. 
Olson,  G.  E. 

Adler,  A.  A. 

Kearny  — 

Sweatt,  C,  H. 

Pines,  S. 
Radio,  H.  M. 

Bock,  B.  A, 
Emery,  H. 

Holbrook,  F.  M. 

Roches  tor—- 
Adams, N.  D, 

Russell,  W.  A. 
Shepparcl,  F.  A, 
Stephenson,  L.  A, 

Asbury  Park  — 
Strevell,  R,  P. 

Lyndhurst  — 

Ehrlich,  M.  W. 

Shovlln— 

Nesdahl,  E,  ^ 

Weiss,  C.  A. 
Wymore,  F.  C. 
Zink,  D.  D. 

Atlantic  City— 

Strouse,  S.  B. 

Maplewood  — 

Evans,  W.  A, 
Kepler,  D.  A, 

St.  Paul- 

Maplewood"— 

Audubon*— 

Kylberg,  V.  C. 
Smith,  M.  S. 

Arnold,  E,  Y, 
Backstrom,  R.  E. 

Walters,  A,  L. 

Sanbern,  E.  N. 

Morchantville  — 

Bamum,  C.  R. 
Buenger,  A, 

St.  Louia— 

Belmar  — 

Binder,  C.  G. 

Fitta,  C,  D, 

Barry,  J,  G,,  Jr. 

Merkel,  F,  P, 

"M       +   l«l 

Gausman,  C,  E* 
GUI,  J.  w. 

Baysc,  H,  V, 
Bradley,  K.  P. 

Bloomfleld— 

Bentz,  H 

Haslett,  H,  M* 
Mickey,  D.  W. 

Carlson,  E,  E, 
Clegg,  R.  R. 

•  Hochuli,  H.  W. 
Tenney,  D. 

Newark- 

Hyde,  L.  L. 
Jones,  E,  F. 
Kolinsky,  M,  D, 
Lindberg,  A.  F. 
McNamara,  W; 
Oberg,  H.  C, 
Penttsi,  R, 
Boucher,  R.  C. 

$ha*w«  H,  W. 

Cooper,  J.  W. 
Corrigan,  J,  A* 
Davis,  C,  R, 
DuBois,  L,  J, 
Edwards,  D.  F, 
Fagin,  D.  J, 
Farvey,  J.  D. 
Fillo.F.  B. 
Foster,  J.  M. 
Gilmore,  L.  A. 
Griffin,  J.  J. 
Groasoiann,  H*  A. 

Camden— 

Brown,  W.  M. 
Kappel,  G.  W.  A. 
Lanning,  E.  K. 
Webster,  E.  K* 
Webster,  W, 
Webster,  W.,  Jr, 

Collinflswood'— 

Plum,  L,  H. 

Alt,  H.  L. 
Ashley,  C,  M, 
Bryant,  P,  J, 
Carey,  P.  C, 
Carrier,  W.  H, 
Day,  V,  S, 
French,  D. 
Holton,  J.  H. 
Ingda,  M, 
Lelnroth,  J.  P, 
Lewis,  L,  L. 
Lewis,  T, 

Wtowtorl&fo,  M*  S, 

Hamilton,  J,  E, 

Ba$t  Oranfte— 

Lvle,  J,  I. 
Matullo,  J.  R. 

tfttfaut1  Oarove— 
^IftiQit*  0*  J. 

Hasten  T*.  j" 
Kent,  J.  K,   \ 
Langenbem  E.  B, 
Maimhan,  J.  & 

Atldnaon,  K.  B, 
Gombm  H,"  B,"  , 

Morehouae,  H.  P, 
Rachal,  J,  M. 
Ray,  L.B. 
Raymfer,  W.  F,,  Jr* 

^>  It^ySfftiti**"*''*^'  '  ' 

McLeary,  H,  W. 

Racki  jS*«  C- 

Rice,  It  B. 

K&dfya*.  C,  W, 

M€M&i|Qiu  T«  W* 

JL^liy,  J*  H» 

Soule,  L.  C. 

Tallmadge  W» 

Stelnmetz,  C,  W,  A. 

W£^  F.D 

Eodei£*l«eVf  G.  B. 

Turno  W'  G*  *W» 
WordSam,  H. 

Tavanlar,  E.  V, 
West,  P, 

AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

North  Arlington— 

Evans,  C.  A. 

New  Rochelle— 

Fife,  G.  D. 

Bermel,  A.  H. 
Vogelbach,  O. 

Farnham,  R. 
Farrar,  C.  W. 
Gifford,  C.  A. 

Abrams,  A. 
Farley,  W.  F. 

Finch,  S.  B. 
Fleisher,  W.  L. 
Flink,  C.  H. 

Orange  — 

Hamlin,  C.  J.,  Jr. 

Rose,  H.  J. 

Frank,  O.  E. 

Crawford,  J.   H.,  Jr. 

Harding,  L.  A. 
Heath,  W.  R. 

New  York  City— 

Frey,  G.  O. 
Friedman,  M. 

Pater  son  — 

Hedley,  P.  S. 

Addams,  H. 

Galloway,  J.  F. 

Cox.  H.  F. 

Hexamer,  H,  D. 

Ames,  C.  F. 

(Brooklyn) 

Pryor  F.  L. 

Jackson,  M.  S. 

Ashley,  E.  E. 

Gautesen,  A. 

Johnson,  E.  E. 

Atherton,  G.  R. 

(Brooklyn) 

Perth  Amboy  — 

Kamman,  A.  R. 

Bachler,  L.  J. 

Genchi,  B. 

Simkin,  M. 

Landers,  J.  J. 
Love,  C.  H. 

Balsam,  C.  P. 
(Brooklyn) 

(Brooklyn) 
Gianm'ni,  A.  A. 

Plainfield— 

Madison,  R.  D, 

Barbera,  H.  A. 

Gilmour,  -A.  B. 

Hedges,  H.  B. 

Mahoney,  D.  J. 

Barbieri,  P.  J. 

Glorc,,  E.  F. 

Tobin,  G.  J. 

McTernan,  F.  J. 

Barnum,  M.  C. 

Goldberg,  M. 

Mosher,  C.  H. 

Baum,  A.  L. 

(Brooklyn) 

Ridgefield  Park- 

Murray,  T.  F. 

Beebe,  F.  E.  W. 

Goldschmidt,  O.  E. 

Davis,  A.  C. 

Rente,  H.  W. 

Bennett,  E.  A. 

Gordon,  P.  B. 

Roebuck,  W.,  Jr. 

Bennitt,  G.  E. 

Gornston,  M,  H. 

Ridgcwood  — 

Shelney,  T. 

Berman,  L.  K, 

Goulding,  W. 

Pitts,  J.  C. 

Snyder,  J.  S. 

Bernharcl,  G 

(Brooklyn) 

Strouse,  S.  W. 

(Lawrence,  L.  I.) 

Haigney,  J.  E, 

Roselle  Park  — 

Karnpish,  N.  S. 

Thornton,  R.  T. 
Voisinet,  W.  E. 

Bernstrom,  B, 
Bilyeu,  W.  F. 

(Brooklyn) 
Hamburger,  F,  G, 

Walker,  E.  R. 

Blackburn,  E.  C.,  Jr. 

Hament,  L. 

South  Orange— 

Wendt,  E.  F. 

(Hempstead,  L.  I.) 

Hartman,  F.  S. 

Hansen,  C.  C. 

Yager,  J.  J. 

Blackmail,  A.  O. 

Hateau,  W.  M. 

Blackmore,  J.  J. 

Heibel,  W.  ET 

Summit  — 

Eggertsville— 

Blackshaw,  J.  L. 

Henry,  A.  S.,  Jr. 

Oaks,  O.  0. 
Teaneck— 

Eiss,  R,  M. 
Hirschman,  W.  F. 

(Brooklyn) 
Bodinger,  J.  H. 
Bolton,  R.  P. 

Herkimcr,  H. 
Herty,  F,  B. 
(Brooklyn) 

Heebner,  W.  M. 

Elmira  — 

Bowles,  P. 
Browne,  A.  L, 

IlerUler,  J.  R. 
(Brooklyn) 

Tenafly  — 

Davis,  B.  C. 

Bruckmann,  J.  C. 

Hiera,  C.  R. 

Redfield,  C. 

McGlenn,  G.  R. 

Buensod,  A,  C. 

(Great  Neck,  L.  I.) 

Bulkeley,  C.  A. 

Hinklc,  E.  C. 

Union  City— 

Geneva  — 

Burbaum,  W.  A, 

t(ilempstead,  L.  I.) 

Taverna,  F.  F. 

Herendecn,  F.  W, 

Callaghan,  P.  F,,  Jr. 

Hmrichaen,  At  F. 

(Brooklyn) 

Hoitman,  C,  S, 

Upper  Montclair  — 

Hamburg  — 

Callahan,  P.  J. 

Holliater,  N,  A. 

Fernald,  H,  B.,  Jr. 

Graham,  C,  H. 

(Great  Kills,  S.  I.) 
Campbell,  F.  B, 

(Brooklyn) 
Hoskina,  H.  L. 

Verona  — 

Stone,  G.  F. 

Hastlngs-on- 
Hudaon—  - 

Campbell,  R.  E, 
Carpenter,  R.  H. 

Hotchklas,  C,  H.  B, 
Howcll,  K.  B. 

West  Orange— 

Reynolds,  T.  W. 

Charles,  T.  J, 
Charlet,  L.  W. 

Hymun,  W.  M. 
IsBcrteU,  H.  CJ. 

Adlam,  T.  N, 
West  New  York— 

Hudson  Falls  — 

HolHster.  E.  W. 

Chase,  C.  L. 
(Brooklyn) 
Cohen,  N. 

Jacobus,  D.  S. 
Janet,  H,  L. 
Johnson,  K.  B. 

Stinard,  R,  L. 

Irvington-on- 

Crone,  T,  E, 
Cucci,  V,  J, 

(W.    New    Brighton, 
S.  I.) 

NEW  YORK 

Hudsott— 

Bastedo,  A.  K. 

Dailey,  J.  A, 
(Astoria,  L.  L) 

Johnston,  W.  H. 
Kagey,  I.  B.^Jr. 

Ithaca- 

Darta,  J*  A". 

Kaufman,  W,  M, 

Albany—- 

Barns, A,  A, 

Daviwn,  R,  L. 

Kenward,  S.  B, 

Anker,  G.  W. 

Sawdon,  W.  M, 

DeBiois,  L.  A. 

(Bay  Shore,  L.  I,) 

Bond,  H.  A, 
Johnson,  H.  S, 

Williams,  J.  W, 
Woods,  E.  H, 

Deely,  J.  J. 
(Brooklyn) 

Keplinger,  W.  L. 
Kimbafi.  D,  D. 

Ryan,  H.  J, 

Denny,  H.  R, 

Knopf,  C. 

Taggart,  R.  C. 

J  a  m  es  town  — 

Donnelly,  R, 

Kuhlmarm,  R. 

Teeling,  G,  A, 

Sharp,  F.  H, 

Downe,  E,  R, 

Ume,  D,  D. 

Blnfthamton—  «•  • 

Kendall— 

Drlacoll,  W.  H, 
Duff,  K, 

(Jackson  Ileightg, 
L  L) 

Brown,  R.  F. 
Martini,  0, 

Stangland,  B,  F, 

Durkee,  M.  E, 
Duryea,  A,  A, 

Lefflngwell,  R.  R. 
Lennon.  J.  O. 

Bronxville  — 

Dornhdra,  G*  A. 

Kenmore—  * 

Candee,  B,  C. 

G  *•*»»«     I     T 

(Belden  Point, 
City  hlftttd) 
Dwyer,  T.  F, 

Leupold,  H.  W. 
Lowy,  M,  R. 
Lucke,  C.  E, 

Buffalo— 

Beman,  M.  C. 
Booth,  C,  A, 
Cherry,  L.  A. 

rcen,  j  .  j  . 
Roseberry,  J,  H. 

Larchmont"— 

Gaylor,  W.  S. 

(Brooklyn) 
Badic,  J,  G, 
Kelts,  H,  B, 
(Brooklyn) 
Elliott,  If, 

Lyle,  E.  T, 
Lyon,  P*  S, 
Ma!  man.  H, 
(GlendaK  L,  L) 
Mandevilk,  E,  W, 

Cheyney,  C\  C, 
Creasy,  R.  E. 
Criqul,  A,  A. 
Currier,  C,  H. 

Lockport— 

Bishop,  C.  R, 
Saunders,  L,  P, 

Engle,  A, 
Everetts,  J.,  Jr. 
Fangler,  P.  E, 
$ay,  F,  C* 

(Brooklyn) 
Mariao,  &  A, 
Markush,  E.  U. 
ManhalL  H*  H, 

Danforth,  N.  L. 
Davis.J. 
Day,  H.  C. 
Dyer,  0,  K, 
Erdie,  G.  F. 

Mt,  Vemon-~ 

Kreitag,  F.  G, 
Northon,  L, 
Obert,  C,  W. 

Fealey,  D.  R. 
Few  man,  A.  M, 
Fenner,  N,  P. 
Ferrero,  H.  L 
Fiedler,  H,  W. 

Martia,  G,  W, 
McClotishan,  C, 
(Brooklytt) 
McKlever,  W,  H. 
MeUitfei,  W»  S, 

48 


ROLL  OF  MEMBERSHIP 

Mehne,  C.  A. 

Sterne,  C.  M. 

McLenegan,  D.  W. 

Richard,  E.  J. 

Mcinke,  H.  G. 

(Long  Island  City) 

Vogel,  A. 

Royer,  E.  B. 

Meisel,  C,  L. 

Stewart,  C.  W. 

Welch,  L.  A.,  Jr. 

Sigmund,  R.  W. 

Meyer,  C.  L. 

Still,  F.  R, 

Smith,  J.  A. 

Meyer,  H.  C.,  Jr. 

Strock,  C. 

Snyder  — 

Sproull,  H.  E. 

Miller,  C.  A. 

Strunin,  J. 

John,  V.  P. 

Washburn,  M.  J. 

Monroe,  M. 

Sutton,  F. 

Winther,  A. 

Montgomery,  0.  C. 
Moore,  R.  E. 
(Brooklyn) 

(Babylon,  L.  I.) 
Syska,  A.  G. 
Thomson,  T.  N. 

Syracuse  — 

Acheson,  A.  R. 
Woese,  C.  F. 

Wright,  K.  A. 
Cleveland  — 

Morse,  F.  W. 
Moss,  E. 

(Huntington,  L.  I.) 
Tiltz,  B.  E. 

Tarrytown— 

Allman,  N.  S. 
Andes,  W. 

(Brooklyn) 

Timmis,  W.  W. 

Abraham,  L. 

Avery,  L.  T. 

M  under,  J.  F.,  Jr. 

Tisnower,  W. 

Weiss,  A.  P. 

Bailey,  E.  P.,  Jr. 

Munier,  L.  L. 
Munro,  E.  A. 

(Brooklyn) 
Torrance,  H. 

Utica— 

Brooks,  F.  W. 
Brueggeraan,  A.  R. 

(Lynbrook,  L.  I.) 

Tucker,  F.  N. 

Steinhorst,  T.  F. 

Cohen,  P. 

Murphy,  C.  G. 
Neale,  L.  L 

(Freeport,  L.  I.) 
Tuach,  W. 

White  Plains- 

Conner,  R.  M. 
Davis,  J.  R. 

Neary,  D.  A. 

(Brooklyn) 

Johnson,  L.  O. 

Dickenson,  F.  R. 

Offner,  A.  J. 

Tyler,  R.  D. 

Zibold,  C.  E. 

Eveleth,  C.  F. 

O'Hare,  G.  W,,  Jr. 
Olsen,  G.  E. 

Vetleaen,  G.  U. 
Vivarttas,  E.  A. 

Williamsville— 

Fonda,  B.  P. 
Fox,  O. 

(Arverne,  L.  1.) 

(Brooklyn) 

Rente,  S.  R. 

Geissbuhler,  J.  O. 

Olvany,  W.  J. 
Osburn,  R.  M, 

Vogt,  J.  H. 
Waechter,  H.  P. 

Yonkers  — 

Gottwald,  C. 
Graham,  W.  D. 

Pabst,  C.  S. 
Patorno,  S.  A.  S. 

(Thompkinsville, 
S.  L) 

Archdeacon,  H.  K. 
Bense,  W.  M. 

Harvey,  L.  C. 
Hendrickson,  J.  J. 

Pfuhler,  J.  L. 

Wachs,  L.  J. 

Brabbe6,  C.  W. 

Kalinsky,  A.  G. 

(W.  New  Brighton, 

(Brooklyn) 

Deutchman,  J. 

Kartorie,  V.  T. 

S.  L) 

Wagner,  F.  H.,  Jr. 

Goerg,  B. 

Kitchen,  F.  A. 

Phillips,  F.  W.,  Jr. 

Walker,  W.  K, 

Hayter,  B. 

Klie,  W. 

(Brooklyn) 

Wallace,  G.  J. 

Kelly,  J.  G. 

Levy,  M.  1. 

Pihlman,  A.  A. 

(Elmhurst,  L.  1.) 

Rainger,  W.  F. 

Martinka,  P»  D. 

Pinder,  P.  H. 

Wallace,  W.  M,,  U 

Stitt,  A.  B. 

Matzen,  H.  B. 

Place,  C.  R. 

(Hollis,  L.  I.) 

Zuhlke,  W.  R. 

O'Gorman,  J.  S.r  Jr. 

Pohlc,  K,  F. 

Walton,  C.  W.,  Jr. 

Miles,  J.  C. 

Presdee,  C.  W. 
Price,  K.  H. 

Waring,  J.  M.  H. 
White,  E.  S. 

NORTH    CAROLINA 

Prendergast,  J.  J, 
Quinlivan,  L.  P. 

(Rivcrhead,  L.  1.) 

Whitdaw,  H,  U 

Rather,  M.  F. 

Purdy,  R.  B. 

Wilder,  K.  L. 

Charlotte  — 

Repko,  J.  J. 

Purinton,  D.  J. 

Wilson,  H.  A.,  Jr, 

Brandt  E.  H. 

Schick,  K.  W. 

Quigley,  W.  J. 
Uuirk,  C.  H. 
Raffes,  A. 

Winquiat,  W.  J. 
(Brooklyn) 
Wohi,  M,  W. 

Christian,  C.  W. 
Hodge,  W.  B. 
Small,  B.  R. 

Schmidt,  R.  H. 
Schurman,  J.  A. 
Sennet,  L.  E. 

Raislcr,  R.  K. 

Spencer,  R.  M, 

Raynis,  T. 

Oriskany  - 

Hi&h  Point- 

Steffner,  E.  F. 

(Woodhaveu  L.  1.) 
Reed,  J.  K. 

Oakey,  W.  K, 

Gray,  W.  E. 

Tuve,  G.  L. 
Vanderhoof,  A.  L. 

Reynolds,  W.  V. 

Ossluin&  — 

Winston  -  Salem  ~— 

WetzelU  H.  E. 

Richardson,  H.  T. 

Hooper,  V.  F. 

Bahnson,  F.  F, 

Wilhelm,  J.  E. 

Riley,  C.  L, 
Riley,  R.  C.          » 
(Jamaica,  L,  I.) 

Patchoque  — 
Jalonack,  I.  G. 

Turner,  M.  E. 
NORTH  DAKOTA 

Wise,  D.  E. 
Cleveland  Height*— 

Ritchie,  E.  J. 

Davis,  R.  G. 

Ritter,  A. 

Pelham  Manor 

"  -i"-'"m-~- 

Rodgers,  F.  A. 

Rodman,  R.  W, 
Rosenberg,  P. 

Peacock,  J.  K. 

Grand  Forks  — 

Pesterfield,  C.  H. 

Columbus- 

Rosa,  J.  0. 

Port  Jorvis—" 

Brown,  A,  L 

Roth,  C,  F. 
Rozettv  W*,  Jr, 

Trimmer,  C.  M, 

OHIO 

Sherman,  R.  A. 
Slayter,  G, 

Ruppert,  E.  H, 

Rochester--* 

-—""•"•-'—' 

Wheeler,  O.  J. 

Schneider,  W,  G, 
Schoepiiia,  P.  H, 
Schulze.  B.  H. 
Scott,  d  E, 

Betlam,  H.  T. 
Coe,  R.  T. 
Cook,  R.  P. 
Hakes,  L.  M. 

Belief  on  taine  — 

Quay,  D.  M. 

Berea— 

Williams,  A,  W. 
Cuyahoga  Falls- 
Humphrey,  D.  E, 

Scott,  G*  M, 
Scribner,  E.  D. 
Seelbach,  H. 
Seelig,  A.  E, 
Sell  man.  N.  T, 

Hutchina,  W.  H. 
Sheldon,  N.  E, 
Stacy,  S.  C, 
Welah,  H,  S. 

Curtii,  H.  F. 

Cincinnati—- 

Bird,  C. 

Dayton- 
Hull,  H,  B, 
LaSalvia,  J,  J. 
Moore,  H.  W. 

Senior,  L  L. 
Seward,  P.  H. 
(Brooklyn) 
Shepard,  E*  C. 

Siebs,  C,  T, 

Rome— 

Lynch,  W.  L, 
Steels,  M.  G. 

\  Breneman,  R.  B. 
Buford,  J.  W. 
Coombc,  J. 
Doyle,  W,  J. 
Floyd,  M. 

Williams,  F,  H. 
East  Cleveland*- 

Morria,  F.  H, 

'      Nobia,  H.  M. 

Scarsdaltt  —« 

Green,  W.  C, 

Read,  R,  R, 

(Brooklyn)  ' 
Skidmore,  J.  G. 
(Long  Island  City) 
Sklemrik,  L,» 
Stack,  F.  C. 
(Flushing,  L.  I.) 
Staples,  W.  H, 
S^srnberg,  E. 

Curnmlng,  R.  W. 
Uliraan,  H*  G. 

Cannon,  C.  N, 
Fauat,  F.  H. 
Harrington,  E,  D. 
Hunssiker,  C,  E. 
James,  J.  W* 

Helburn,  L  B. 
Houliston,  G,  B. 
Hust,  C.  E, 
Kiefer,  C.  J, 
Kitchdl,  H.  N. 
Kramlg,  R,  E.,  Jr, 
Hllen,  H.  A, 
Pistler,  W.  C, 
Powera,  L.  G. 

Stark,  W.  E. 
ElyrJa— 

Maynard,  J.  E. 
Lakewood  — 
I*ongcoy,  G,  B, 
Ramsey,  R.  F, 
Vermere,  E.  J. 

AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 

Mansfield  — 

Beech  wood,  Del.  Co.~ 

Philadelphia- 

Wilmot,  C.  S. 

Yardley,  R.  W. 

Kipe,  J.  M. 

Adams,  B. 
Bartlett,  C.  E, 

Woolston,  A.  H. 
Wright,  M.  B. 

Newark  — 

Bradford  — 

Black,  H.  G. 

Pittsburgh  — 

Simpson,  D.  C. 

Black,  W.  B. 

Blankin,  M.  F. 

Aston,  J. 

Norwood  — 

'Braun,  J.  J. 
Motz,  O.  W. 

Butler  — 

Karges,  L. 

Bogaty,  H.  S. 
Bornemann,  W.  A. 
Braemer,  W.  G.  R. 
Breitenbach,  G.  C. 

Beighel,  H.  A. 
Blackmorc,  G.  C. 
Blackmore,  J.  S. 

Caldwell,  A.  C. 

Brauer,  R, 

Painesville  — 

Ghambersb  ur  &  — 

Cassell,  j.  D. 

Burns,  J.  R. 

Hobbs,  J.  C. 

Kottcamp,  H.  A, 

Clarkson,  R.  C.,  Jr. 
Clodfelter,  J.  L. 

Bums,  R. 
Bushnell,  C.  D. 

Shaker  Heights  — 

Bolz,  H.  A. 
Gary,  E.  B, 

Steubenville  — 

Smith,  R.  H. 

Cheltenham  — 

McElgin,  J.  W. 

Cynwyd  — 

Smith,  W.  F. 

Cornell,  J.  C. 
Culbert,  W.  P. 
Dambly,  A.  E. 
Davidson,  L.  C. 
Davidson,  P.  L. 
Donovan,  W.  J, 
Eakins  'W' 

Carr,  M.  L. 
Cheeseman,  E.  W. 
Collins,  J.  F.  S.,  Jr. 
Comstock,  G.  M. 
Dice,  E.  vS. 
Dorfun,  M.  I. 
Edwards,  P.  A. 

Toledo- 

E.  Pittsburgh  — 

Eggly,  H.  J.,  Jr. 

Evans,  E.  C. 
Frit?    (7    V 

Baker,  H.  C. 

Goodwin,  W.  C. 

Eichberg,  W.  R, 

Giles!  A!  F.' 

Myers,  F.  L. 
Treadway,  Q. 

Youngstown  — 

Boucherle,  H.  N. 
Chomn,  C.  C. 

Elizabethtown  — 

Dibble,  S.  E. 

Erie- 
Joyce,  H.  B. 

Elliot,  E. 
Engel,  E. 
Erickson,  H.  H. 
Faltcnbacher,  H.  J. 
Kamiletti,  A,  R. 
Galligan,  A.  B, 
Gant,  H.  P. 

Gunther,  F.  A. 
Hecht,  F.  II  . 
Heilman,  R.  H. 
Hickey,  J.  W. 
Houghten,  F.  C. 
Humphreys,  C.  M. 
Huettner,  1L  F. 

OKLAHOMA 

Glenolden  — 

Grossman,  H.  E, 

Harrisburg.  — 

Eicher,  H.  C. 
Geiger,  L  H. 
Lutz,  J.  H,,  Jr. 

Oilman,  F.'  W. 
Glaasey,  J.  W. 
Hackett,  H.  B. 
Hibbs,  F.  C. 
Hires,  J.  E. 
Hoft,  P.  J. 
Hunger,  R.  F. 

Ingold,  J,  W, 
Kellner,  D*  C. 
Kennedy,  O,  A. 
Kennedy,  P,  V. 
Lofte,  J.  A. 
LuU,  W.  J. 
Muclilinjs,  L.  S. 
Msxjjinii  P.  1^« 

Bartlesvllle— 

Yates,  G.  L. 

Norman  — 

Bowman,  J.  W. 
Chestnutt.  N.  P, 
Cook,  A.  B. 
Dawson,  E.  F. 
Frarapton,  A.  C. 
Husky,  S.  T. 
McKinley,  C.  B. 

Haverford— 

Black,  E.  N.,  3rd 

Jenkln  to  wn— 

Slight,  I. 

Hynca,  L,  P. 
Ickeringill,  J.  C, 
Jellett,  S.  A. 
Kclble,  F,  R. 
Kriebel,  A.  E. 
Leopold,  C.  S. 
Lewis,  G.  C. 

McGinncjBS,  J.  E. 
MeDonagle,  A. 
McGuigan,  L.  A. 
Mdntush,  F.  C, 
McMunn,  A.  H. 
Miller,  K.  A. 
Nn."W   A    V 

Rauh,  E.  M. 
Sneed,  R.  B. 
Sonney,  K.  J. 

Johnstown  — 
Novotney,  T.  A. 

MacDade,  A.  H, 
Mann,  L.  B. 
Marks,  A.  A. 

Nicholla,  P. 
Nordheiu\cr,  C,  L, 
O'Neill  P, 

Oklahoma  City- 

Stitt,  E.  W. 

Martocello,  J.  A. 
Mather,  H,  II. 

Oatcrle!  W.  H. 

Beard,  E.  L. 
Dolan,  R.  G. 
Dugger,  E.  R. 

Kingston  — 

MacDonald,  D.  B. 

McClintock,  A.,  Jr. 
Mellon,  J.  T,  J. 
Menaing,  F.  D, 

Pierce,'  W.  MacL. 
Pittock,  L.  B. 
Reed   I   G 

Emmons,  N.  L. 
Gray,  B.  W. 
Hewlett,  I.  G, 
Loeffler,  F.  X, 
Loemer,  L,,  Jr, 
Miller,  B,  R, 
Miner,  M.  H, 
Rathbun,  P.  W. 
Holland,  S.  L. 
Tauson,  P.  0* 
Tiller,  L. 

Tulsa— 

Burke,  W.  J, 
Jones,  E. 

Lancaster  — 
e'nea,  A. 
oyd,  E.  C, 
Long,  D,  R. 

Lan  sdo  wne  — 

Hansen,  C.  J. 
James,  H.  R, 

Manheim— 

Weltzel,  P.  H, 

MaryavUle— 

Gault,  G.  W, 

Meyer,  J,  W, 
Monday,  C.  E. 
Morgan,  R.  C, 
Morris,  A.  M, 
Naylor,  C,  L, 
Neabitt,  A.  J. 
Neabitt,  J.  J. 
Newcornb,  L.  B. 
Nusbaum,  L, 
Flaw,  C.  W. 
Plewea,  S.  E, 
Powers,  E,  C. 
Pryibil,  P,  L. 
Redstone,  A,  L, 
Rellly,  C.  E. 

Reed!  V,  A. 
Richmond,  J, 
Riddle,  K.  L, 
Rlesmeyer,  E,  H.,  Jr. 
Rockwell,  T,  R 
Seanlon,  K.  S. 
Smyers,  1C,  C. 
Speller,  K,  N. 
Stanger,  R,  B, 
Steen,  J,  M, 
Steggall,  H.  B, 
Stevenaon,  W»  W. 
Strauch,  P.  C, 
Tennant,  R»  J,  J, 
Tower,  E,  S, 

OREGON 

McKeesport"— 

Duuan  T  M 

Ret  tew.  H,  F. 
Rhea,  C.  A. 
Roberts,  H.  L, 

Waters,  G,  G, 
Wright,  C  E, 

Rugart,  1C, 

Corvallis— 

Meftdvllle— 

Sabin,  E,  R. 

Marty,  E(  0. 

WiHey,  E,  C, 

Williams,  L,  E, 

Shanklin.  A.  P. 
ShelHer,  M. 

Smith,  J,  D, 

PENNSYLVANIA 

Merlon- 
Atkins,  T,  J, 

Speckman,  C.  H, 
Stevens,  J,  M. 
Thornburg,  H.  A, 
Tiramls,  P. 

Luck,  A,  W, 
.Nicely,  J.  E, 
Rldtey  X»*irk— 

Allen  town—* 

Korn,  C.  B. 

New  Castle- 
Andrews,  G.  H. 

Touton,  R,  D. 
Traugott.  M, 

Mawby,  P, 
Stalb,  J,  G* 

Ardmore— 

Haynes,  C,  V, 

Sonneborn,  C. 
Norristown—  * 

Trump,  C.  C, 
Tuckerman,  G*  E. 
Walsh,  J.  A, 

Rutledfte— 

Vroom«,  A,  E. 

Jteaver  Falls-— 

Bolsinger,  R,  C. 

Wandi«u«,  F.  W* 

jS^gnUktOXI'***' 

Vfcji  Atan,  W,  T. 

Hucker,  J,  H. 

Wegraann,  A. 

Sbnyer,  H»  E* 

50 


ROLL  OF 

MEMBERSHIP 

State  College- 

TEXAS 

Portsmouth  — 

Ft.  Atkinson  — 

Queer,  E.  R. 

Stubbs,  W.  C. 

Shodron,  J.  G. 

Stroudsburg,  — 

Amarillo  — 

Richmond  — 

Kohler— 

Kiefcr,  E.  J.,  Jr. 

Burnett,  E.  S. 

Carle,  W.  E. 

Hvoslef,  F.  W. 

Tamaqua  — 

Hadesty,  A.  L.,  Jr, 
Koch,  H.  O. 

College  Station  — 

Badgett,  W.  H. 
Giesecke,  F.  E. 
Smith,  E.  G. 

JcJhnston,  J.  A. 
Livingston,  B.  B. 
Pelouze,  H.  L.,  2nd 
Schulz,  H.  I. 
Snyder.  A.  K. 

Kohler,  W.  J.,  Jr. 

LaCrosse  — 

Anderegg,  R.  H, 

Miller,  M.  W. 

Upper^Darby  — 

Eastman,  C.  B. 
Taliaferro,  R.  R. 

Dallas  — 

Bock,  I.  I. 
Boruch,  E.  R. 

University  — 

Peebles,  J.  K.,  Jr. 

Trane,  R.  N. 
Madison  — 

Villanova  — 
Ban-,  G.  W, 

Durning,  E.  H. 
Gay,  L.  M. 
Landauer,  L.  L, 

WASHINGTON 

Dean,  C.  L. 
Hess,  D.  K. 
Larson,  G.  L. 

Carey,  J.  A. 
Wallin&ford— 

Lynn,  J,  H. 
Meffert,  G.  H. 
Witmer,  C.  N. 

Kent— 

Boyker,  R.  O. 

Nelson,  D.  W. 
Plaenert,  A.  B. 
White,  J.  C. 

Arnold,  R.  S, 

Ft.  Worth  — 

Pullman  — 

Menomonie  — 

Wilkes-Barre— 

Skinner,  H.  W. 

O'Connell,  P.  M. 

Schoenoff,  A.  E. 

Santee,  H.  C. 

Houston  — 

Seattle  — 

Milwaukee  — 

Wilkinsburg— 

Groseclose,  J.  B. 

Beggs,  W.  E. 

Berghoefer,  V.  A. 

Campbell,  T.  F, 
Vernier,  M.  G. 

Kiesling,  J.  A. 
Renouf,  E.  P. 

Bouillon,  L. 
Cook,  H.  A. 

Bowers,  A.  F, 
Bowers,  R.  C. 

T 

Cox,  W.  W. 

Brown,  W.  H. 

WWiamsport— 

Irving  — 

Daly,  C.  P. 

Burch,  L.  A. 

Pfeiffer,  J.  F, 

Moler,  W,  H. 

Dudley,  W.  L. 

Elliott,  N.  B. 

Wor  mlcy  sb  ur  ft  — 

Miller  T.  G. 

Kin&ssviile~— 

Richtmann,  W.  M. 

Eastwood,  E.  O. 
Forsyth,  A.  T. 
Cranston,  R.  0. 

Ellis,  H,  W. 
Freeman,  A.  M. 
Hanley,  E.  V. 

York- 
Baker,  L  C. 

San  Antonio—- 
Diver, M,  L. 
Ebert,  W,  A. 

Griffin,  D.  C. 
Hauan,  M.  J. 
Hendrickson,  H,  M. 
MacLeod,  K.  F. 

Haupt,  H.  F. 
Herrmann,  H.  C. 
Jackson,  C.  H. 
Jepertinger,  R.  C. 

RHODE  ISLAND 

Sherman  — 

Mallis,  W. 
May,  C*  W. 

Jones,  E.  A, 
Jung,  J.  S, 

,  J.__,,Jtll,m-.J, 

Reynolds,  J.  A, 

Musgrave,  M.  N. 

Juttner,  O.  J. 

Providence— 

UTAH 

Peterson,  S.  D. 
Pollard,  A.  L. 

Miller,  C.  W. 
Miller,  H.  M. 

Coleman,  J.  B. 

ji_.^...>. 

Twist,  C.  F. 

Noll,  W.  F. 

Gibba,  E,  W. 
Hartwril,  J.  C. 
McLaughlin,  J.  D. 
Moulder,  A.  W, 

Salt  Lake  City  — 

Richardson,  II.  G. 

Zokell,  C.  G. 
Tacoma  — 

Spofforth,  W. 

Peters,  H,  H. 
Randolph,  C.  H. 
Rice,  C.  J, 
Shawlin,  W.  C, 

SOUTH  CAROLINA 

VERMONT 

Yakima— 

McCune,  B.  V. 

Spence,  M,  R. 
Spielmann,  H.  J. 
SwJsher,  S.  G.,  Jr, 

——-—'— 

Burlington  — 

Szekely,  E. 

Caaey,  11.  F. 

Lanou,  J.  E. 
Raine,  J,  J, 

WEST  VIRGINIA 

Volk,  J.  H. 
Wagner,  A,  M. 

Clemaon  Colle&e— 

North  Fentisburft— 

1     Charleston  — 

W?lso"'W.'  H! 

Shenk,  D.  H, 

Breckenrldge,  L,  P. 

Shanklin,  J.  A. 

Racine  —  • 

TENNESSEE 

VIRGINIA 

Thompson,  D, 
Titus,  M.  S. 

Dixon,  A.  G, 

Thomas,  N.  A. 

Memphis— 

Danville- 

Larftent™— 

Donnelly,  J,  A, 

Superior  — 

Campbell,  A.  Q,,  Jr, 

Farley,  W.  S, 

Innis,  H,  R. 

Waite,  H. 

D'Imor,  B.  J. 
Hoshaii,  R,  H. 
Perkins,  R.  C, 

Lynchburg  — 
Doering.  F,  L* 
Wiley,  E.  C. 

WISCONSIN 

Wauwatosa  — 

Page,  H.  W. 

Nashville™"" 

H  rowxi  F« 

Norfolk— 

Ashland  — 

West  Allls— 

Jarrett,  P.  R. 

Nowitsky,  H,  S. 

Trulson,  A,  F. 

Erickson,  M.  E. 

51 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


FOREIGN  COUNTRIES 


AUSTRALIA 

Eaton,  W.  G.  M. 

ENGLAND 

Ellis,  F.  E. 

Sydney  — 

Duncan,  J.  R. 
Hunt,  N.  P. 
Sands,  C.  C. 

Flanagan,  E.  T. 
Gaby,  F.  A. 
Gurney,  E.  H. 
Harrington,  C. 
Heard,  R.  G. 

Buckinghamshire— 

Russell,  J.  N. 
Leeds  — 

BELGIUM 

Henion,  H.  D. 

Jennins,  H.  H. 

Hills,  A.  H. 
Hopper,  G.  H. 
Jenney,  H.  B. 
Leitch,  A.  S. 

London  — 

Bailey,  W.  M. 
Butt,  R.  E.  W. 

Brussels  — 

Mautsch,  R. 

CANADA 

MacKenzie,  J.  J. 
Marriner,  J.  M.  S. 

Greenland,  S.  F. 
Haden,  G.  N. 

Brandon,  Man.  — 

McDonald,  T. 
McHenry,  R.  W. 

Herring,  E. 
Nobbs,  W.  W. 

Yates,  J.  E. 

Millar,  R.  J, 
Moore  H  S 

Middlesex- 

Brockville,  Ont.  — 

O'Neill,  J.'  W. 

Case,  W.  G. 

Davenport,  R.  F. 

Paterson,  J.  S. 

Chipperfield,  W,  H 

Calgary,  Alberta- 

Paul,  D.  I. 
Philip,  W. 

Stockport  — 

Clarke,  S.  S. 

Playt'air,  G.  A. 

Webb,  J.  W. 

Walker.  A. 

Price,  D.  O. 

Edmonton,   Alberta- 

Purdy,  A.  K. 
Ritchie,  A.  G. 

Sutton  — 

Casperd,  H.  W.  H. 

Kelly,  H. 

Roth,  H.  R. 

Shears,  M.  W. 

Swinton  — 

Gait  — 

Sheppard,  W.  G.  F. 

Yates,  W. 

Oke,  W.  C. 
Sheldon,  W.  D.,  Jr. 

Thomas,  M.  F. 
Waldon,  C.  D. 

Trowbrid&e  — 

Halifax,  N.  S.— 

Watson,  M.  B. 
Whittall,  E.  T, 

Haden,  W.  N. 

Eagar,  R.  F. 

Woollard,  M.  S. 

Westminster  — 

Hamilton,  Ont.— 

Wood,  J.  S. 

Faber,  Dr.  O. 

Best,  M.  W. 

Vancouver,  B.  C.~~ 

Wolvorhampton—  • 

Fitzsimons,  J.  P. 

Johnston,  R.  E. 

Tyson,  W.  H. 

Maddux,  O.  L. 

Leek,  W, 

Pennock,  W.  B. 

Libby,  R.  S. 

FRANCE 

McCreery,  H.  J. 

Islington,  Ont.-  — 

Wilson,  G.  T. 

Victoria,  B.  C.  — 

Dijon- 

Kitchener,  Out.  — 

Sheret,  A. 

Bur,  J.  R.  C, 

Beavers,  G.  R. 

Wellington,  Out.™— 

Lille— 

Montreal,  P.  O.  — 

Johnston,  H.  D. 

Neu»  H.  J.  E. 

Darling,  A.  B, 

Winnipeg  Man.— 

Lyon  — 

Fogarty,  Q.  A. 

Glass,  W. 

Goenaga,  R.  C. 

Friedman,  F.  J. 

Jones,  B.  G. 

Garneau,  L. 

Kirk,  C.  D. 

Paris— 

Givin,  A.  W. 

Leonard,  J.  H. 

Beaurrienne.  A, 

Johnson,  C.  W. 

Michie,  D,  F, 

Downe,  H,  S, 

McGrail,  T,  K, 

Steele,  J,  B. 

Modiano,  R. 

Osborne,  G.  H. 

Summers,  E.  T, 

Nessi,  A, 

Phipps,  F.  G, 

Turland,  C.  H, 

Schmuts!,  J. 

Wiggs,  G.  L. 

Montreal,  W.,  P»  Q.— 

CHINA 

GERMANY 

tinton,  J.  P.  - 

Nanklnft— 

Berlin— 

Ottawa,  Ont,— 

Loo,  P.  V, 

Brand!,  O,  H, 

Colclough,  O.  T* 
Gray,  G.  A. 

Shanghai— 

Stutt&art— 

Ke&ina,  Sask.  — 
Stewart,  J.  C, 

Carter,  D, 
Doughty,  C.  J. 
Hart-  Baker,  H*  W, 

Klein,  A, 
HOLLAND 

Kwan,  1.  K. 

Three  Rivers,  Que.— 
Germain,  Q, 
Toronto,  Ont.— 

Loh,  N,  S. 
Merritt,  C,  J. 
Morrison,  C,  B. 
Waung,  T.  F, 

Amsterdam  — 

Overton,  S,  H. 

Allsop,  R.  P. 
Angus,  H.  H. 

CZECHOSLOVAKIA 

INDIA 

Anthe&  L  JL 

Arrowsinith,  J»  O, 

Praa— 

New  Delhi- 

Bfrrell,  A.  L. 
Blackball,  W.  R. 

Brust,  0. 

Heard,  J,  A.  E. 

Boddington,  W,  P, 
Church,  H,  J, 

DENMARK 

IRELAND 

Cole,  G*  E. 
Dickey,  A,  J. 

Copenhagen*— 

Cork- 

Duncan,  W,  A, 

Reck,  W,  K> 

Barry,  P.  I. 

ITALY 

Milan— 

Donzelli,  E, 
Gini,  A. 
Hauss,  C.  F. 

JAPAN 


Osaka — 

Fukui,  K. 
Tokyo*— 

Kitaura,  S. 
Kozu,  T. 
iSaito,  i>. 
yekido,  K, 

MEXICO 


Lomas  de 

Chapul  tepee  — 

Gilfrin,  G.  F. 
Mexico,  D.  P.— 

Darby,  M.  11. 
Martinez,  J.  J. 

NEW  ZEALAND 

Ghristchurch — 

Taylor,  K.  M, 
Vale,  H,  A,  L, 

Dimedln — 

Daviea,  G.  W. 

NORWAY 

Oslo- 
All  sen,  N. 
TjeralunU,  A, 

PHILLIHNK 
ISLANDS 

Moloiloilo— 

Piion,  D.,  jr. 

SOUTH  AFRICA 

J  ohannesb  ur &  -  • 

Carrier,  E.  G. 

SWEDEN 


Ro«ell,  A,  F, 
Stockholm- 

Gille,  H. 
TheorelU  H,  G,  T. 


,  M.  L, 
SCOTLAND 

Dundee-— 

Knoxf  J,  E. 


PAST  OFFICERS 
AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS 


1894 

President Edward  P.  Bates 

1st  Vice-President Wm.  M.  Mackay 

$nd  Vice- President Wiltsie  F.  Wolfe 

3rd  Vice-president Chas.  S.  Onderdonk 

Treasurer Judson  A.  Goodrich 

Secretary L.  H.  Hart 

Board  of  Managers 

Chairman,  Fred  P.  Smith 
Henry  Adams  A,  A.  Gary 

Hugh  J.  tturron  James  A.  Harding 

'  "    "          ~  L.  H.  " 


Edward  P.  Hates,  Prcs. 


.  Hart,  Secy. 


Council 


Chair  man »  R,  C.  Carpenter 
Albert  A.  Cry^r  Chas.  W.  Newton 

F,  W*  Foster  Ulysses  G.  Scollay,  Secy. 


1897 

President Wm.  M.  Mackay 

1st  Vice-President H.  D.  Crane 

8nd  Vice-President Henry  Adams 

Srd  Vice-President A.  E.  Kenrick 

Treasurer Judson  A.  Goodrich 

Secretary H.  M.  Swetland 

Board  of  Managers 

Chairman,  R.  C.  Carpenter 
Edward  P.  Bates  Stewart  A.  Jellett 

W.  S.  Hadaway,  Jr.  Wiltsie  F.  Wolfe 

Wm.  M.  Mackay,  Pres.        H.  M.  Swetland,  Secy. 


John  A.  Fish 
Wm.  McMannis 


Council 

Chairman,  Albert  A.  Cryer 


James  Mackay 
B.  F.  Stangland 


1895 

President......  , ..........................Stewart  A.  Jellett 

1st  Vic,f -President...,, Wm.  M.  Mackay 

Snd  Vice-president. ,...Chaa.  S,  Onderdonk 

3rd  Vice-President ......D.  M,  Quay 

Treasurer............ , Judson  A.  Goodrich 

Secretary.., .,.. „ L,  H.  Hart 

Board  of  Managers 

Chairman,  Jaraea  A,  Harding 
Ceo.  B,  Cobb  Ulysses  G.  Scollay 

Wm.  McMannis  B.  F.  Stangland 

Stewart  A,  Jellett,  Pres.  L.  H,  Hart,  Secy. 

Council 

Chairman,  R,  C.  Carpenter 
Henry  Adama  T,  J.  Waters 

Edward  P.  Batei  Albert  A,  Cryer,  S««y« 


1898 

President........ Wiltsie  F.  Wolfe 

1st  Vice-President T.  H.  Kinealy 

Snd  Vice-President A.  E.  Kenrick 

Srd  Vice-President.^ ...John  A.  Fish 

Treasurer. Judson  A.  Goodrich 

Secretary. ...Stewart  A.  Jellett 

Board  of  Managers 

Chairman,  Wmu  M.  Mackay 
Thomas  Barwice  A.  C,  Mott 

John  A,  Connolly  Francis  A.  Williams 

Wlltalc  F.  Wolfe.  Pres.       Stewart  A.  Jeliett,  Secy* 

Council 

Chairman,  R.  C,  Carpenter 
Henry  Adams  W»  S,  Hadaway,  Jr. 

Albert  A*  Cryer  Wm,  McMannis 

Wllttle  F,  Wolfe,  Fm.       Stewart  A,  Jeilett,  Secy. 


..........  .................  ..»,  ..........  ...R.  C.  Carpenter 

Xst  Vice-president,  ......  ......,«.....,....,.......,.D.  M,  Quay 

Viti-Pr*sident»mt..»f»  ........  .....Edward  P*  Bates 


$rd 


»mt..»f»  ........  .....  * 

.»~  .......  .............  .....  ....F,  W.  Foster 

.,„  Jfudion  A,  Goodrich 
.M.^..»M......<»»L.  H,  Hart 


Board  of  Managers 

Chairman,  Wm.  M.  Mackay 
Hugh  J.  Barron  Stewart  A*  Jellett 

W/k  Hadaway,  Jr.  Wilt*  F,  Wolfe 

E,  C,  Carpenttr»  Pfe*»  U  H,  Hart,  Seey. 

Council 

Chairman,  A,  A,  Cary 

Albtrt  A,  Cry«r  B*  F.  Staaglautd 

W».  McMnnnls  J.  J» 


1899 

Presidtint,..^  ......  ...„.,...»„.„........  ........  .....Henry  Adams 

1st  Vic^PresMent.^..^  ........  ..,..„.  ...........  D,  M.  Quay 

&nd  Vice-President.^..,  ...............  .  .........  A,  E,  Kenrick 

9fd  Vice-president  ........  .,  ........  ....Francis  A,  Williams 

.^  ............  „.„  .............  Judson  A.  Goodrich 

.,  ......  ......  ..............  ».,......Wft*  M.  Mackay 


Board  of  Managers 

Chairman,  Stewart  A,  Jellett 
B.  H.  Carpenter  Wm,  Kent 

A*  A.  Cary  Wtttale  F,  Wolfe 

Htnry  Adams,  Prttt  Wm.  M.  Mackay,  Secy, 

Council 

Chairman,  R»  C.  Carpenter 
John  Gormly  Wm,  McMarmU 

W.  8»  Hadaway,  Jr,  B,  F,  Stangland 

Henry  Adam«t  jPrw*          Wm.  M,  Mackay,  Secy. 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


1900 

President „ D.  M.  Quay 

1st  Vice-President A.  E.  Kenrick 

2nd  Vice-President Francis  A.  Williams 

Treasurer Judson  A.  Goodrich 

Secretary Win.  M.  Mackay 

Board  of  Governors 

Chairman,  D.  M.  Quay 
Wm.  Kent,  Vice-Chm.         D.  M.  Nesbit 
R.  C.  Carpenter  C.  B.  J.  Snyder 

John  Gorrnly  Wm.  M.  Mackay,  Secy, 


1901 

President J.  H.  Kinealy 

1st  Vice-president A.  E.  Kenrick 

2nd  Vice-President^ Andrew  Harvey 

Treasurer Judson  A.  Goodrich 

Secretary. Wm.  M.  Mackay 

Board  of  Governors 

Chairman,  J.  H.  Kinealy 
Wm,  Kent,  Vice-Chm.         John  Gormly 
R,  C.  Carpenter  C.  B.  J.  Snyder 

R.  P.  Bolton  Wm.  M.  Mackay,  Secy. 


1902 

President A.  E,  Kenrick 

1st  V^ce- President Andrew  Harvey 

Snd  Vice-President Robert  C.  Clarkson 

Treasurer. , Judson  A,  Goodrich 

Secretary ..Wm.  M,  Mackay 

Board  of  Governors 

Chairman,  A.  E.  Kenrick 
John  Gormly,  Vice-Chm.    J,  H.  Kinealy 
R.  C,  Carpenter  C,  B.  J.  Snyder 

Wm.  Kent  Wm,  M.  Mackay,  Secy. 


1905 

President Wm.  Kent 

1st  Vice- President R.  P.  Bolton 

Snd  Vice-President C.  B.  J.  Snyder 

Treasurer Ulysses  G.  Scollay 

Secretary Wm.  M.  Mackay 

Board  of  Governors 

Chairman,  Wm.  Kent 
R.  P.  Bolton  Tames  Mackay 

C.  B.  J.  Snyder  B.  F,  Stangland 

B.  H.  Carpenter  J.  C,  F.  Trachsel 

A,  B.  Franklin  Wm.  M.  Mackay,  Secy* 


1906 

President , John  Gormly 

1st  Vice-President C.  B.  J.  Snyder 

2nd  Vice-President T.  J.  Waters 

Treasurer Ulysses  G.  Scollay 

Secretary , Wm.  M,  Mackay 

Board  of  Governors 

Chairman,  John  Gonnly 
C.  B.  J.  Snyder,  Vice-Chm.  Jamea  Mackay 
R.  C.  Carpenter  B.  F.  Stangland 

Frank  K.  Chew  T,  J.  Waters 

A.  B.  Franklin  Wm,  M.  Mackay,  Secy* 


1907 

President,.....,. , ,C.  B.  J.  Snyder 

1st  Vice-President....,,.....,.,......,., James  Mackay 

2nd  Vice-president ......,....Wm.  G,  Snow 

Treasurer..... .„„.„ Ulysses  G,  Scollay 

Secretary ........  Wm.  M,  Muckay 

Board  of  Governors 

Chairman,  C.  B.  J,  Snyder 
Jamea  Mackay,  Vice-Chm^  Frank  K.  Chew 
R.  E.  Atkinson  A.  B.  Franklin 

R.  C.  Carpenter  Wm.  G.  Snow 

Edmund  F,  Capron  Wm.  M.  Mackay,  Stcy. 


1903 

President H,  0.  Crane 

IstV^e-Presidmt Wm.  Kent 

£nd  Vice-President R.  p.  Bolton 

Treasurer Judaon  A.  Goodrich 

Secretary.^...,.......,. ...........Wm.  M.  Mackay 

Board  of  Governors 

*  •«  *  ^       Chairman,  H.  D,  Crane 

C.  B.  J.  Snyder,  Vie+Chm*  A,  E.  Kenrick 

R,  C,  Carpenter  Geo.  Mehring 

John  Gonnly  Wm,  M.  Mackay,  Secy, 


1908 

President  —  .............  .....  ......  .  .......  ,  .........  James  Mackay 

1st  Vict-President  ..........  ,......„„..,...  .  Jas.  D,  Hoffman 

$nd  Vice-President.^  ____  ..........  ......  ...,  B.  F.  Stangland 

Treasurer  .................  .....  ........  ...  .......  Ulysses  G,  Stollay 

Swyda/y..........................  .......  ......  ...Wm.  M.  Mackay 

Board  of  Governors 

Chairman,  James  Maekay 
Jas,  D.Hoff  man,  Vic*-Chm.  John  F,  Hale 
B,  F.  Stangland  August  Kchm 

R.  C,  Carpenter  C,  B,  j, 

Frank  K.  Chew  Wm, 


,  j,  Snyder 
,  M.  Mackay, 


1904 

„.»,.„. Andrew  Harvey 

-•-•  -  ,v  -•""-- — ...John  Gormly 

end  Vic^President .....Robert  C.  Clarkson 

Treasurer ; .........Ulysses  G.  Scollay 

Secretary., ...,.,.Wm.  M.  Mackay 

Board  of  Governors 

Chairman,  Andrew  Harvey 
John  Gormly  H.  D.  Crane 

Robert  C.  Clarkson  A,  E.  Kenrick 

J.  J,  Biackmore  C*  B,  1  Snyder 

R.  C.  Carp«ntttr  Wm,  M.  Mackay,  S#ey. 


1909 


President.. 

1st  Wc^^«*irf^.I..7.7/,...Z. 

Snd  ""     "     '" 

Secretary.* "...".". 


...,.Wm.  G,  Snow 

August  Kehm 

.........B*  5.  Harriioii 

I.  Mackay 


Board  of  Governors 


Chairman, 

August  K«hm»  Vice-Ckm. 
John  R,  AUen 
R.  C»  Carpfent«r 
B.  S.  H&riiio» 


G, 

Samuel  R.  L«wl§ 
Jfttn«»  Mactoy 
B»  F.  &ta,$dam<! 
Wia*  M.  Mwfcity, 


ROLL  OF  MEMBERSHIP 


1910 

President Jas.  D.  Hoffman 

1st  Vice-President R.  P.  Bolton 

£nd  Vice-President.^ Samuel  R.  Lewis 

Treasurer Ulysses  G.  Scollay 

Secretary..,..,.. , Wm.  M.  Mackay 

Board  of  Governors 

Chairman,  Jas.  D.  Hoffman. 
R.  P.  Bolton,  Vice-Chm.  John  F.  Hale 
•Geo.  W.  Barr  Samuel  R.  Lewis 

R.  C.  Carpenter  James  Mackay 

Judson  A.  Goodrich  Wm,  M.  Mackay,  Secy. 


1911 

President..,. R,  P.  Bolton 

1st  V+CG-Presidcnt John  R.  Allen 

£nd  V ice-President..., A.  B,  Franklin 

Treasurer ..Ulysses  G.  Scollay 

Secretary Wm.  W,  Macon 

Board  of  Governors 

Chairman,  R.  P.  Bolton 
John  R.  Allen,  Vice-Chm.     A.  B,  Franklin 
John  T.  Bradley  Jas.  D.  Hoffman 

R.  C.  Carpenter  August  Ivehm 

James  H.  Davis  Wm,  W.  Macon,  Secy. 


1912 

President „ , John  R.  Allen 

1st  Vice-President^ John  F.  Hale 

Snd  Vice-President Edmund  F.  Capron 

Treasurer , James  A.  Donnelly 

Secretary .....,..., „ Wm.  W,  Macon 

Board  of  Governors 

Chairman,  John  R.  Allen 

John  F,  Hale,  Vice-Chm.      D  wight  D,  Kimball 
Edmund  F.  Capron  Samuel  R,  Lewis 

R.  P,  Bolton  Wm.  M.  Mackay 

Jas.  D,  Hoffman  Wm.  W«  Macon,  Secy. 


1913 

,,.*.,*,,,,..., John  F.  Hale 

1st  Vice-president ,...,...A.  B,  Franklin 

&nd  Vice-President ..Edmund  F.  Capron 

Treasurer, ».,», ..„.„., ...James  A.  Donnelly 

Secretary,*,, „....„„......,...... Edwin  A.  Scott 

Board  of  Governors 

Chairman,  John  F,  Hale 

A.  B.  Franklin,  VicihChm,  James  A.  Donnelly 

John  R,  Allen  Dwight  D,  Kimball 

Edmund  F,  Capron  Wm.  W,  Macon 

R,  P,  Bolton  lames  H.  Starmard 

Frank  T»  Chapman  Theodore  Wdnahank 

Ralph  Collamore  Edwin  A,  Scott ,  Secy, 


£nd 


1914    » 
,M...WM,M»«. .Samud  R,  Lewli 

^.................Edmund  F.  Capron 

L.; ..... Dwight  D,  Kimball 

„...,.„.»..*...,  Jamta  A,  Donnelly 
,««.,*,,...»,.».,.„,„ J.  J,  Blackmore 


Council 

Chairman.  Samuel  R.  Lewis 


John  IL  Alien 
Frank  T*  Chapman 
Frank  I*  Cooper 
,  Donnelly 


1915 

President Dwight  D.  Kimball 

1st  Vice-President Harry  M.  Hart 

Snd  Vice-President Frank  T.  Chapman 

Treasurer Homer  Addama 

Secretary J.  J.  Blackmore 

Council 

Chairman,  Dwight  D.  Kimball 

Harry  M.  Hart,  Vice-Chm.  Samuel  R.  Lewis 

Homer  Addams  Frank  G.  McCann 

Frank  T.  Chapman  J.  T.  J.  Mellon 

Frank  I.  Cooper  Henry  C.  Meyer,  Jr. 

E.  Vernon  Hill  Arthur  K.  Ohmes    ' 

Wm.  M.  Kingsbury  J.  J.  Blackmore,  Secy. 


1916 

President Harry  M.  Hart 

1st  Vice-President Frank  T.  Chapman 

Snd  Vice-President Arthur  K.  Ohmes 

Treasurer.^ Homer  Addams 

Secretary Casin  W.  Obert 

Council 

Chairman,  Harry  M,  Hart 
F.  T.  Chapman,  Vice-Chm,  Dwight  D.  Kimball 
Homer  Addams  Henry  C.  Meyer,  Jr, 

Charles  R.  Bishop  Arthur  K,  Ohmes 

1  Frank  I,  Cooper  Fred  R.  Still 

Milton  W.  Franklin  Walter  S,  Timmis 

E.  Vernon  Hill  Casin  W.  Obert,  Secy. 


1917 

President , J.  Irvine  Lyle 

1st  Vice-President, Arthur  K.  Ohmes 

Snd  Vice-President Fred  R.  Still 

Treasurer,, * -.*...» Homer  Addama 

Secretary.., , Casin  W.  Obert 

Council 

Chairman*  J.  Irvine  Lyle 

A.  K.  Ohmee,  Vice+Chm.  Harry  M.  Hart 

Homer  Addama  E.  Vernon  Hill 

Davis  S.  Boyden  Jamee  M.  Stannard 

Bert  C,  Davia  Fred  R.  Still 

Milton  W.  Franklin  Walter  S.  Timmia 

Charles  A.  Fuller  Casin  W.  Obert,  Secy. 


President.~~~,..* « 


J8wd!  Wc*-PywW*w/» 

Tr«fljttr^. 

5«<T«fafy.... 


1918 

„..„ ............Fred  R.  Still 

..... Walter  S*  Timmis 

,.,., E.  Vernon  Hill 

.Homer  Addams 
.Ctain  W,  Obert 


Council 

Chairman,  Fred  R,  Still 

W.  S.  Timmis,  Viw-Ckm*  J,  Irvine  Lyle 

Homer  Addams  «  E,  Vernon  Hill 

William  H,  Drtacoll  Frank  G,  Phegley 

Howard  H,  Fielding  Fred.  W,  Power* 

»,  Oant  Charaplain  k,  Eilty 

W.  Kiwbatt  Ca«ta  W.  Obert»  5#c 


AMERICAN  SOCIETY  of  HEATING  and  VENTILATING  ENGINEERS  GUIDE,  1935 


1919 

President Walter  S.  Timmis 

1st  Vice-President E.  Vernon  Hill 

2nd  Vice-President Milton  W.  Franklin 

Treasurer.^ Homer  Addams 

Secretary Casin  W.  Obert 


1923 

President H.  P.  Gant 

1st  Vice-President Homer  Addams 

2nd  Vice-President E,  E.  McNair 

Treasurer Wm.  H.  Driscoll 

Secretary C,  W.  Obert 


Council 

Chairman,  Walter  S.  Timmis 

E.  Vernon  Hill,  Vice-Chm.  Frank  G.  Phegley 

Homer  Addaras  Fred.  W.  Powers 

Howard  H.  Fielding  Robt.  W.  Pryor,  Jr. 

Milton  W.  Franklin  Champlain  L.  Riley 

Harry  E.  Gerrish  Fred  R.  Still 

George  B.  Nichols  Casin  W.  Obert,  Secy. 


Council 

Chairman,  H.  P.  Gant 

Homer  Addams,  Vice-Chm.  E.  S,  Hallett 

W.  H.  Carrier  Alfred  Kellogg 

T.  A.  Cutler  Thornton  Lexvis 

S.  E.  Dibble  E.  E.  McNair 

Wm.  H.  Driscoll  Perry  West 

Casin  W.  Obert,  Secy. 


1920 

President.. E.  Vernon  Hill 

1st  Vice-president Champlain  L.  Riley 

2nd  Vice-President Jay  R.  McColl 

Treasurer,^ Homer  Addams 

Secretary Casin  W.  Obert 


1924 

President Homer  Addams 

1st  Vice-President , S.  E.  Dibble 

2nd  Vice-President William  H.  Driacoll 

Treasurer Perry  West 

Secretary..... F.  C.  Houghten 


Council 


Chairman,  E.  Vernon  Hill 
C.  L.  Riley,  Vice-Chm.         Jay  R.  McColl 
Homer  Addams  Ge<         "" 

Jos.  A,  Cutler 
Wm.  H.  Driscoll 
A.  C,  Edgar 
Alfred  Kellogg 


:orge  B.  Nichols 
Robt.  W.  Pryor,  Jr. 
W.  S.  Timmis 
Perry  West 
Casin  W.  Obert,  Secy. 


Council 

Chairman.  Homer  Addams 

S,  E.  Dibble,  Vice-Chm,  W.  E.  Gillham 

F.  Paul  Anderson  L,  A.  Harding 

W.  H,  Carrier  Alfred  Kellogg 

J.  A.  Cutler  Thornton  Lewis 

William  H.  Driacoll  Perry  Wcat 

H.  P.  Gant  F.  C.  Houghten,  Secy. 


1921 

President Champlain  L,  Riley 

1st  Vice~Pr«sident Jay  R.  McColl 

Snd  Vice-President H.  P.  Gant 

Treasurer.^,,*,, „., » Homer  Addams 

Secretary. Casin  W.  Obert 


1925 

President .,, .  S.  1C,  Dibble 

1st  Vice-President. „. Wm.  H.  Driacoll 

£nd  Vice-President ...,R  Paul  Anderson 

Treasurer. , , .„....,..„.....,.., /Perry  West 

Secretary.., .....F.  C.  Houghten 


Council 

Chairman,  Champlain  L,  Riley 

Jay  R.  McColl,  Wct-Chm.  E,  S.  Hallett 

Homer  Addams  E.  Vernon  HU1 

Joa.  A.  Cutler  Alfred  Kellogg 

Samuel  E.  Dibble  E.  E.  McNair 

Wm.  H*  Driscoll  Perry  West 

H.  P,  Gant  Caste  W.  Obert,  Secy. 


Council 

Chairman,  S,  E,  Dibble 

Wm*  H.  DriicoIl.PYce-CViw,  W.  T*  Jones 

Homer  Addams  Thornton  Lewis 

F,  Paul  Anderson  J.  H.  Walker 

W,  H.  Carrier  Ferry  West 

T,  A,  Cutler  A.  C.  Willmrd 

W.  E.  GiJlham  F.  C.  Houghten,  .9, 


, ...Jay  R,  McColl 

Ut  Vice-President^.., „..„.,....„ H,  P,  Gant 

$nd  Vi6*.J*rHident». ...Samuel  E*  Dibble 

TY#asur$F ..................................... .„.„„. Homer  Addama 

Secretary Ctsin  W.  Obert 


.,............«.» «,.,«....,.W.  H*  Driscoll 

1st  Vie*-Pr*ati&<tnt.~ ,,,,,F*  Paul  Anderson 

SSnd  Vfa-Prtsidtmt*^....*.- .,..A,  C  WIHwd 

Treasurer, „„.„„„„...«... ...,.W«  E,  GHIham 

.,...,...M.«...«...«..»..A.  V. 


Council 

Chairman,  Jay  R.  McColl 
H,  P,  Gattt,  Vice»Chm,        L.  A,  Harding 

Homer  Addama  E.  E-  McNair 

Jos.  A.  Cutler  H,  J.  Meyer 

Samud  E,  Dibble  C,  JU  Riley 

Wm,  H.  0ri8coll  Perry  West 

E.  S.  Haitett  Casin  W.  Obert,  , 


Council 

Chairman,  W.  H*  Drfccoll 
F.  Paul  Anderson,  Vica»Chm*      C>  V.  Hayne* 
W.  H.  Carrier  W.  T.  Jone« 

J*  A,  Cutl«r  E.  B,  Lutnaenber 

S.  E,  Dibble  Thomtoa  Lewlt 

W.  E.  Gillham  J.  F*  I     " 

A,  C,  Wmard 


ROLL  OF  MEMBERSHIP 


1927 

President F.  Paul  Anderson 

1st  Vice-President , A.  C.  Willard 

8nd  Vice-President Thornton  Lewis 

Treasurer W.  E.  Gillham 

Secretary A.  V.  Hutchinson 


Council 

Chairman,  F.  Paul  Anderson 


A.  C.  Willard,  Vice-Chm. 
H.  H.  Angus 
W.  H.  Carrier 
W.  H.  Driscoll 
Roswell  Farnham 
H.  H.  Fielding 
W.  E.  Gillham 
C.  V.  Haynes 


John  Howatt 
W.  T.  Jones 
J.  J.  Kissick 

E.  B.  Langenberg 
Thornton  Lewis 
J.  F.  Mclntire 

H.  Lee  Moore 

F.  B.  Rowley 


1928 


President A.  C.  Willard 

1st  Vice- President Thornton  Lewis 

8nd  Vice-President L.  A.  Harding 

Treasurer.^ W.  E.  Gillham 

Secretary A.  V.  Hutchinson 

Council 

Chairman*  A.  C.  WUlard 
Thornton  Lewis,  Vice-Chm,         C.  V.  Haynes 
F.  Paul  Anderson  John  Howatt 

H.  H,  Angus  W,  T.  Jones 

W,  H.  Carrier  J.  J.  Kissick 

E.  B.  3 


N.  W.  Downes 


Langenberg 


Roswell  Farnham  J.  F.  Mclntire 

H,  Lee 


W.  E.  Gillham 


F.  B.  Rowley 


-ee  Moore 


1929 

President,. ,. Thornton  Lewis 

1st  Vicf- President. -.-L,  A.  Harding 

$nd  Vice-President.., .W,  H,  Carrier 

Treasurer W.  K.  Gillham 

Stcrttary'  [.L.I......... .—A,  V,  Hutchinson 

Technical  Secretary.... ....,.,.— ............ P.  D.  Close 

Council 

Chairman,  Thornton  Lewis 
L,  A.  Harding,  Vice-Chm.  John  Howatt 

H,  H.  An«ui  w.  T.  Jones 

W.  H.  Carrier  K.  B,  Langenberg 

N.  W,  Downee  G.  L.  Larson 

Roswell  Farnham  K.  C.  Mcintosh 

W.  E,  Gillham  W.  A,  Rowe 

C.  V,  Haynes  F.  B,  Rowley 

A.  C.  Willard 


1930 


^,,  ......... 

1st  Vice-President  ................  ,  ....... 

$nd  Vice-President.^  _______  ........  . 


.. 
Secretary  „  ...... 


1931 

President W.  H.  Carrier 

1st  Vice-President F.  B.  Rowley 

2nd  Vice- President W.  T.  Jones 

Treasurer F.  D.  Mensing 

Secretary _"..." "." ." A.  V.  Hutchinson 

Technical  Secretary P.  D.  Close 

Council 

Chairman,  W.  H. 
F.  B.  Rowley,  Vice-Chm. 

D.  S.  Boyden 

E.  K.  Campbell 
R.  H.  Carpenter 
J.  D.  Cassell 

E.  O.  Eastwood 
Roswell  Farnham 
E.  H.  Gurney 


..t,  A.  Harding 
-.W,  H,  Carrier 
....F.  B,  Rowley 


;••"•„•  , 

..»..».  ......  A.  V.  Hutchlnson 

.  ____  .  ...........  ,..,,«,.P.  D.  Close 


Council 

Chairman*  L.  A.  Harding 

W,  H*  Can-tor,  Vte+Chm.  John  Howatt 

H,  H.  Angua  W*  T.  Jones 

D,  S.  Boyden  E.  B,  Langenberg 

R,  H.  Carpenter  G.  L,  Larson 

J,  D.  Ca*s*ll  Thornton  Lewis 

k  W,  Downs*  F.  C,  Mclntosh 

Rotwell  Farnham  W.  A,  &ow« 

C,  W,  Farrar  F,  B.  Rowley 


Carrier 
L.  A.  Harding 
John  Howatt 
W.  T.  Jones 

E.  B.  Langenberg 
G.  L.  Larson 

F.  C.  Mclntosh 
F.  D.  Mensing 
W.  A.  Rowe 


1932 

President F.  B.  Rowley 

1st  Vice-President W.  T.  Jones 

&nd  Vice-President C.  V.  Haynea 

Treasurer.^ , » ....F.  D.  Mensing 

Secretary A.  V,  Hutchinson 

Technical  Secretary P.  D.  Close 

Council 

Chairman,  F.  B.  Rowley 

W.  T.  Jones,  Vice-Chm.  F.  E.  Giesecke 

D.  S.  Boyden  E.   Holt  Gurney 

E.  1C  Campbell  C.  V.  Haynes 
R.  H.  Carpenter  John  Howatt 
W.  H.  Carrier  G.  L.  Larson 
John  D.  Cassell  J.  F.  Mclntire 
E,  0.  Eastwood  F.  D.  Mensing 
Roswell  Farnham  W.  E.  Stark 


1933 

President W.  T.  Jones 

1st  Vice-President.^.* ,C.  V.  Haynes 

$nd  Vice-president John  Howatt 

Treasurer..., D.  S.  Boyden 

Secretary,.,., , A.  V.  Hutchinson 

Council 

Chairman,  W.  T.  Jones 

C.  V,  Haynea,  Vice-Chm,  G.  L.  Larson 
R,  H.  Carpenter  J.  F.  Mclntire 
F.  D.  Caasell  W.  E.  Stark 

D.  S.  Boyden  E.  K.  Campbell 
John  Howatt  E.  0.  Eastwood 
k  C.  Mclntosh  R.  Farnham 

L*  W.  Moon  E,  H,  Gurney 

F.  E.  Giesecke  F.  B.  Rowley 


1934 

President , ..... ..C.  V,  Haynea 

1st  Vice-President ,... John  Howatt 

find  Vict-President ........G.  L,  Larson 

„„.,.     „ ....D.  S.  Boyden 

A.  V.  HutchinBon 


Council 

Chairman,  C.  V.  Haynes 

John  Howatt,  Vics-Chm.  W,  T\  Jones 

M.  C.  Beman  O.  L.  Larson 

D  S,  Boyden  J-  F,  Mclntire 

Albert  Buenger  F.  C.  Mclntosh 

R.  H.  Carpenter  L.  Walter  Moon 

LD,  Cas«ell  O.  W,  Ott 

F,  E.  Gie«ecke  W.  A.  Russell 

E.  H,  Curney  W.  E,  Stark 


57 


1 34  974