See other formats

Full text of "American Society Of Heating And Ventilating Engineers Guide 1935 Volume 13"

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. 



COPYRIGHT S 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 follows 1 : 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, O 2 20.75, water vapor 1.2, A 0.93, CO 2 0.03 and H 2 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^ p 2 , p s , etc. = the pressure of the gases or vapors corresponding to 
the observed temperature, then 

P = pi + pz + p 3 , 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 - t 1 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. 
e 1 - 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, 
C p , and that at constant volume, C v , 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 range 1 
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 Am a 

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 





0.08636 


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 Ara a 

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 





0.08625 


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 P 2 represent the initial and final absolute pressures, and V\ and 
F 2 represent corresponding volumes of the same mass, say one pound of 

V P 
gas, then =? = --, or PI FI = P 2 F 2 , 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 



5 

Ii 

H 

c^ 

s 

2 

Q 



CM Q\ Os ON *> X s * 

* i CO *O vOO^f 

*HO\t^ to fO i 

-i>- vo o o o ^o 

ooo ooo 




& = 



3^-5 



?^ 



S 1 



ooo ooo ooo ooo 



OOO OOO OOO 000 OOO 000 OOO 



3 O3 CM CM C 



000 000 OOO 000 OOO 000 OOO 



ooo ooo ooo ooo ooo ooo ooo 



i IVOrH THCO 1 ^-! 

t- t CO i I to to 

- rh csj ON vO CO 
!>**' f-^>.*>. j>. \O *O ^O vo vO \O ^O to to to 



OOO OOO OOO 000 OOO 000 OOO 




ooo ooo ooo ooo ooo ooo ooo 



OOO 000 000 000 OOO OOO 000 



OOO^O 
OOC^to 
t-\OTt< 

t~*. J>. t< J^ Jf<*. V 



CO i-l O 
iO Q <M 

CN O r*> 

383 



ooo ooo ooo ooo ooo ooo ooo 



OO 00 ir^vo^O VQtotO * rf CO CO CO CNJ CM Csj -rH T-I -rH 



OOO OOO OOO 000 OOO OOO OOO 



VO vO IO tO to to to O to \O 




000 000 OOO 000 000 OOO 000 



O 00 O ^-O\C4 OO <N t^ VO i-l OO CO 

OfOOO (Mt^fO 00-^Os -^ Qs CS tO 

lOCO'r-l OOO*> IO^*C^ i < Q\ OO VO .. 

t^t-J>. t-^. O y? "O O VQ O to to to to to O ** ^K Th PO 

0*00 0*0*0* 000* O*0*O 0*OO* OOO O'0*O" 



s 



a 



II 
B 



as 



s 

'3 I 



o . 



| ft 

fe rt o, 

in 



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. 

Charles 1 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, F c , 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 
Tables 2 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, e t , at the dry- 
bulb temperature, or the ratio of the actual density, 8, of the vapor to 
the density of saturated vapor, 8 t , 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, W t 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 



CMOCTHIOPO TH TH O VO C 
tO CMTH Q TH - ~ ~ ~ 

CO -^ O vO CM 



OO T# TH 1>- 



f^ O CO TH CM coco 

CO IO OO -^ CM CO 00 OO O CM "* t-~ TH VO TH t *<* CM O 

TH CO *O co TH ON t TH vO O ""T OO CO t-- CM VO TH vO TH 



O-ICMCMCO 



EH < F" 

i> m 
pi &. 

^ 



CM VO t ON IO O C 



S^ 
. .^ 



_.-.,,O r-vOto^tO VOOOTHT^OO 

^^ O CO VO O CM TT lOt^-ONTHCO tot^OCM-^ 



i CM CM CM CM 



) 00 TH ro to CO O cOXOf-OCM ri< M3 

< CM vHON-T-tOT^ CMOOOf-^iO CO TH \O O 

< ON "* OO CO 00 CO COCOr-CMr- CMt~*O\CM 



'-i'^-i CM CM CO fO rf 



OOCOCOONOs OS ON O O O 



4 



OMO O LO O vOTHt>CMCO CO ON ^ O IO TH ^-. ON CM to OO TH ^ t* O CO VO ON CM 
iO\o^-*>.OO 00 ON ON O O TH TH CM co CO r^ -^ rf to to to vO vO vO t-- X^- J>- J>- OO 



T-( ^-i T-H TH *-< -rH *-H -^-1 CM CM CM CM CM CM CM CM CM CM CM CM CM CM CM C<I CM CM CM CM CM 



OOCOOOCOOO CO 00 Tf ON ^ ON 
lOvOvOt-^t-- OO 00 ON ON O O 



"*< ON -HH ON^ C 



^3! ^55^^^ S>ar*fi 



vo vO vo 



H TH T-I CM CM CM CM CM CM CM CM CM CM CM CM CM CM CM CM CM CM CM CM 



iO O OO ON "H 



88 88888 88888 8 




3CM CM CM CO COCO 



^ VOCOON TH ro 
-r^CMOO O TH rj< 
?1>.OOQNO THCMCO 




B 

^2-g 
^1^ 



CM ON OO-rHvOtOoO 



-COT- i O'-HC 

----- ^ 



VO'^-i t^-voiot-^o 

CM t>- -rH VO -r-l VO CM 
i-t --H CM CM CO CO T*I 



vOv 
O 



. - . . - O CM to ON 

"* =32S^^ 



ON 
O 



to CO . - 

ON O rH CM CM co TJ4 to vO l>- 
-rH CM CM CM CM CM CM CM CM CM CM 



KS 

111 

11! 



OCM**VOOO OCMrhvooo OCM-^VO 



OO O 

CM COC 



10 



CHAPTER 1 FUNDAMENTALS OF HEATING AND AIR CONDITIONING 



ON O> O CM to ONrJ<Ot IO 
iO O vO TH vO TH r^- co OO -^ 



CM CO 10 TH CO to CS ON VO CO TH ON t- to co T 



5 ON \O rj< ro CM co iO OO TH O CO O ON CO ON CM 

- ON CN to oo i i Tti r*- O -* r^- --H 10 co CN vO 1-1 10 O r 



r*. oo co oo ON C\ ON O O O *HI -i CS CM ?r> P 



g; 

81 

$* 

H 



> CS THt- ON CM < 



O *-< v-i TH t~i CM 






iO CO i-i rt< r- Ocor^-OrO vOOfOOO fCOOrOt^ 
CO CO O\ ON ON OOO-^-t-<-H t-i <M C^J CS T> CO CO <# ^ ** 



CO ro fO <T) CO 



CM <M CS CS CM CS CM CS O4 CS CO CO CO ^O CO CO CO CO fO CO CO fO rO CO rO CO CO ro CO 



TH Tin co co ON t-- cs TI co ON co to to co O 
CSONt^iO'* -^ to VO CO O TFOQCOONVO CO CM CS CM CO 



CO CO CO CO ONONONOO ^-t T-I TH cs CM 




OtOi-HCOvO to vO CO TH vO coOOi-HTj* CO -^ CS TH CO 

-- 10 ON cs vo O ^ oo co t^ cs >> CM -t"- cs J> cO ON to TH 

~ ~ O O TH TH TH CS CM CO CO ^^ "^ to IO VO NO !> OO 

OC3 OOOOO OOOOO OOOOO 



CO O TH COON tO \ 

TH \OTH vOCM ON v, 

THcs-*ito^ co<; 



cococococo rH rt< rj ri* rft 



10 vo vo O vo vo r 



COOOCOCOON 



CMO CO t-.t- 




.8888 88888 

O ' ' ' ' O 



h 

**s 

l 



g-s 

11 



?33 



CO O ^5 ^4* CS 
CO CO CO CO CO 



piocs rHc 

TH < 

cs ^ 

CD 



5^-coto oo c 



K 00 Ov vOCMvOt-vO 

5 ON CO CO-<*O*--iO 

5OOTH COVOONTHTH 

>vot^ J>. t- *> OO CO 



11 



AMERICAN SOCIETY of HEATING and VENTILATING ENGINEERS GUIDE, 1935 



S 3 li 

SrS-s* 

<3E*8f! 

sllfs 

SBM5 



5 TH Tf O Ox TH vO CO TH OO kO VO TH OO Q to *O ON *>- OO to *>- 
3 ON OC-acovOOO TH -^ r TH to O TH O to TH t^-THTHON-t>" 



TH TH cq VO CM THC 
CO TH ON t- VO to 



I 8 



iOvO 00 O\ O - i ro 
CO PO CO CO Tti T*< ** 



O 
IO 



^ M 

35 



THiOOxTHOO CNvOOTHOx CO J>. TH 
THCOiOOOO CO to 00 O CM lOt^O 



OOOOOOOOO\ 



CM CS CS CM CN CSC 



O 
10 



QfOiOOOO cotOOOOco vOCO^-fCOvO OO^-HCOVOOO T-H rt< VD O\ ^ 
'OvO'O'Or- t>- JO- f- OO OO OOOOOsOsOs Os O O O O ^ *H TH TH CNJ 

co CO CO co co co co CO CO CO CO CO CO co CO 



OOCSJJ>-^cO 



vO Ox -^ I-H -^-l CO OO O to OO CO CS -^ Ox f 



v> TH O cs oo TftOt^coO t^- T** CS 
lOvOOt^r^ OO O\ Ox O t-l TH Ofl CO 

-r-l T-4 -rH iH fH *-t ^H ->-! CS CS C^I CNI CS C 




vOvOt^OstN vo O O M O Ov OO OO O CO vo TH i>- rH cr<> CM CM to OO OJ OO vo to to vo 
THi>-OcOt- OTTl^"THto OOcNvOTHto ON TH 00 CO OO co 00 CO OO TH ON to TH i>. CO 



O\C3\OOO ^"^ 1 



OO OO Ox O\ O ^ -^ 




S8888 88888 88888 



1! 

12 

y 

85 
1| 
II 

aZ 
l 

I- 8 

a3 o 
^1 

S-s 



ON^O\-Sc 

THT$T?THTH toiOiOtoiO 



i i 00 *-i CM Ox - 
O THO^Ooa 



3 VO vot^t^- t^-t 

d 



oo c 
O 



^^ 

O O TH T-i TH" 



12 



CHAPTER 1 FUNDAMENTALS OF HEATING AND AIR CONDITIONING 



Cv O r O O O 
r^LOiO \O 00 i-< 



Y-iXDCSQOiO CO t 



? rOCS TH PO i-t C 



t'-.f*. oo oooo 



O\ i-< Tf <O p> CS t 
O\OOOO -! C 



"5 C\ **- O O oo r- CS OO i/i O 00 
00 -H iO ON <*O O\ r O Os * CS 



f-vj ro to !> Os TH to to t < 



J s 



-icotoooo roJ>-rocsto CST^CSO\*>- I-HNOTH 

* .t-- 1> t OO OOOvTHCOtO OO T- I to O VO fO CS lO 

THTHTH TH CS CS CS CO ^ 



-i fO vo 
VOCS *^ 



fc fa^O 

3S| 



5 CS *O O to O ro r>- < < O O TH CS ro to \O t* OO O ^-t CS fO t/> O t-* O 

500OfOto t^OCStot^- OCSTfvOOO c5 CS T** >O Q\ THfOVOt-.O rip 

LOtotovOvO vOOt^t^t^ t^.odcX3CCOO O\ O r4 CS fO vOVOr^OOOs -< CS PO rf iO OQ 

CNCSCSCSCS cscscscscs cscscscscs csrocOfOfO rocorororo T# * ^< ^ ^ -^ 





I>.OOOOONO -^ ^ 



CSO\fOrOr^ ^ Q CO Os lO r*< *-f -H* Ov Q ^ 

to o\ o ^ oo ^o O t- O r- ooiooooo cs 

OvOf-COOO OvOi-iPOT^ 'OONfr>Jr^v> r* 

,H ,-, ^ ,*"-< THCSCStCNCN CMCMfO^O^ t^ 



^OO>TH^ vo Ov TH Tt* O ONCS*<^r-O fS to OO Q CO *O O CS rf t^- O CO ^O OO H t- 
CSCSCSCOCO ro CO ri< ^ Ti< ^* vo to VO *O Ol>.OOOTH CS CO O VO f-* O O -( CS rj* <O 



10 CO CNJ ^O -r 
- 



3Jo ! i i ! 
^cs j i : : 



T) CM O\ ^O rj< ro CS *- CS ro 

o<sjcoor- os'^-<fOot^ 

LOiOiOlOtO mvOOvOvO 

OOOOO OOOOO 



cs*-n^t^ ^*<or>-o\to ro 



- 
OOOO OOfH 



OCS TH OOt CS 
r OO TH to C?\ to 
THTH THCSCSCSCO 



\|>.IOOO to 



COf-IOOVOro T-t 



THTHO\\OTH 



8S 



^StOOeo 
\ lO CO TH TH 



O ^-r fO ^ y 



>0000 000 



cs-* o\ io 

SSS^s 



3 

I 

3 



TH-^OOCSOQ T^THOOkO O 
OtOVDOPO t-THiOOOC- - 
THi-tTHCSCS CSPOcOtO^ 



\cstooo cs^cst>c 






O * O\ * 

CS O CS 



li 



1 



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 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 P er P und of d T air - 

* 0618 lb per P Und f dry air ' 

The water vapor added per pound of dry air must be (W z - 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. 
W 2 = 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 ^-^Q 1 587) = - 01245 lb P er P und of ^ ain 

w, = 0.622 (2992^042) " - 00887 lb P er p und of dr y 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 out 3 that the equation for the process of adiabatic saturation, and 
hence for a process of constant wet-bulb temperature, is: 

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

and using c Pa = 0.24 and c Ps = 0.45 

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

where 

h*f s latent heat of vaporization at t 1 , 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 + c Ps (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 













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 t l , 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, 

S r = 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 F 
as the datum for dry air, and the saturated liquid at 32 F as the datum 
for the water vapor) is 

h = c Pa (t - 0) 4- W h s = 0.24 (* - 0) + W h s (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. 
7f s = 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 h s 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 W z (1059.2 + 0.45 fc) - W l (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 = S 3 - Si (16) 

where 

2 3 = 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, 

E h + 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 CHART 4 

The Bulkeley Psychrometric Chart 5 , 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, 1931 V 
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 
(W 3 ~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 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. L c 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>Ci=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 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 (H 2 0) 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 


2V 2 "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. Prss. 


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 


Vf K 


Vg 


hf 


hfg 


fcg 


Sf 


Sfg 


s g 


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 





0.0522 


925.Q 





925.0 


1.0785 





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. 

e r = 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. 
e 1 = 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, U t 5.8; at a velocity of 800 fpm, U t = 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 = W t = 0.01578 Ib (Table 5). Satura- 
tion pressure of the vapor at 70 F = e t = 0.73S6 in. Hg. 
From Equation 7 r 

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 , = 0.072886 Ib. 

Lo,t A 

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

A A7oafV7 ~ 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, 

Cp a = 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, 
CP a = 0.24. 
A f f g = 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, P 2 , 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) 

P 2 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 

h s = 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 researches 1 , 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 
observation 2 . 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 decreased 3 , probably by an inhibitory 7 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 
temperature 4 , 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 \VoRK a 

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 





0.2 


225,000 


0.1 


7 


0.6 


80 


96.7 


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 tract 5 . These are considered to be the potent 
factors in the increased susceptibility to gastro-intestinal disorders in hot 
summer weather. The vie f im 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, 
Moss 6 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). 

6 Some 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 medulla 7 . The extra output of adrenin hastens 
heat production which protects the organism against cooling. Bast 8 
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 England 9 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 industries 10 . 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). 

8 St 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). 

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

M Iron 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 trait 11 . 

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 body 12 . 

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 tests 13 ' 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. Laboratory 16 in the U. S. Bureau of 
Mines, Pittsburgh, and later studies at the Harvard School of Public 
Health 17 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&ndiiiomng t 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 Laboratory 19 . 




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 period 20 . 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 Society 21 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 radiation 22 . 

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 Health 17 . 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 Howell 23 , Miura 24 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- 
orders 25 - 26 - 27 . 

For the premature infant, a high relative humidity of about 65 per cent 
is demonstrably beneficial to health and growth 28 , and according to 
Huntingdon 29 , 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 Nose t 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 elsewhere 30 . 

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 Laboratory 32 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 
occupants 22 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 STANDARDS 33 

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 Baetjer 34 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 work 2 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. IV r 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. 

Experiments 35 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 



raoo. I . | - ! : ' . : i i i 


' ' i ; * i i K^l j 1 1 i ; Pi 






! ' ' ' "i 'X: i ( ! ; i T 




1200 -) t i : i . 1 i i ' 


i ' j ' ' , ' \ ' 












- ! ! - ' i ' , ; ' r \ 




u. I tOO' i 1 i ! [ 1 i ' t i ! M 


-H-^-r ' . ! ' ' ' : \ | f 




u -U.,, ,.;. .. ^_- 




















j .. _ | \ 




v3 - "'v I 4-- i 


\ 


1 j 


__5zi _ ^ . 


IV 






. i , i ' T ) .. 






L J - - i \1 


i 




^L^s --}--- | |\j 






1 "^"^ " ! V 




fiArt ' \ |Sy^ 


M - i" 1 ^ s V 






} S rt Px 1\ 






" r ' " >-v \ \ 




oc __jk ^ 


I > \ 




> ._ . J... i>^ 


i<^ >. M 










gfiftfi 


r i X TV 




^ 600 : ::__|_ +_ . 


- i ' i 1 1 i 1 1( * H ' \ i\ ; 






TT^ TT+'VN 






i \ 
















cc i U 1 :__ . 


i N -x .ft} ' 


ffS -- 


400 r n L -- 




H:::: 






jrjj 








H 3OO IJI_-. 


1 1 ' i ' 


A J_ L 




._ U 


PW" = 






_^2 j 






5t L 






^WH- 


g :-:-:::-:- 




==3ff 


100 




----K 








A ,: ,: 


L 1 


::::lt 



3O^ 4<f SCf 6<r^ 7O 80P 90 100" 
EFFECTIVE TEMPERATURE "FAHR. 

FIG. 6. RELATION BETWEEN TOTAL HEAT Loss FROM 

THE HUMAN BODY AND EFFECTIVE TEMPERATURE 

FOR STILL AiR a 

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? Jf 1 ^ 6 " 11 ? 1 Exchanges Between the Bodies of Men Working and the Atmospheric Environment, by 
N 2 M h CI lVn 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 



laoo 
p MOO 



IOOO 



* 8OO 

700 

6OO 

. SCO 

s 

400 

rr 300 
d 

i 200 

H 100 



30 



4<f 53 s 60^ 7tf 80T 90 
DRY BULB TEMPERATURE FAHR. 



100 



FIG. 7. RELATION BETWEEN SENSIBLE HEAT Loss 

FROM THE HUMAN BODY AND DRY-BULB TEMPERATURE 

FOR STILL AiR a 

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 































P 6 


p' 000 H- 




























| JJ 


E 
























" y 1 ^ ' 


































E * 
























" y 








, i....... 






















s 






C 2 


of j j 

O 8-GOp~ 




















~~yt 


f " 




""!""/ 


S. J 






















/ 






/* X 


" 


1 ~ 


















/ 








/ 


5000 o 






























































Uj ..j, 
























,yr i / 


../ 


-" > 


<? 1 














/ 










7*-V^ 


/( 


. < 


tf 1 














/] 








/ 


>/-~ / 




&40(^> cc 


> 1 


j , ! 






L-J- 




=? 






_P 




V 


/~-*-*f > i 


-L_*_- 


~ C 




! i ! 










/ 










^ V 


-' / 1 








! ] - i-1 










r' 


j 






V 


^ 


















f 














/[ j 








! j, l-l 








f 












X 1 






u-3000 *; 




j ! 






^ 














i4^ 


c j 




Ui 


fld U- 


. j J j | 






4 




-1 




,1 


/" ' 


y + 


/ 




! 


p OL 


31 -L. 


u. U t 




4-i/ 








, _, 






"7*" 




| " |"'- ' 


-pr 


"~ ffi 




-' n i 




I/ 












i 


^Lr* 










UJ 3OO 


-H-4fr 


' j/ 








-y; 










!- 1 






^,3XDO h- 





M Mj 1 


"7TT 




-t 








^ 








1 -[y* i 


!j 


- S 
































j^S ^^* U 






















L^- 


^j t" 












x 






^i 








* 






f J 






1 ' ' j. 




^ 








i 








>T-| 








i I00 ff 




^_ 






\ 


1 






^ 




i ! 






E z 


fe ' 


- r - i 


' ' 






-- 


-r~ 


i 




-4 




i 




1 ! 


w S 


P - 


. i i ! . 


! 1 








\ 




















3CT 4i 


O u 


e 


rf 




6 


tf* 




7 


CT 


8 


Cf 9 


J 


u 



DRY BULB TEMPERATURE 



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 Laboratory 37 . 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 F T 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 air f 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 / Q QQ ^ 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 CO 2 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 CO 2 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*SO t 


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 LIBRARIES 1 

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 schedules 2 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 F (assumed saturated) = 0.85 
Btu per pound at 49.85 F saturated = 20.11 

Recirculated air = ^5 38 ~ 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 wind 1 . 

Forces due to Stack Effect 2 

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). 

2 See 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 velocity 3 , this resistance becomes a 
limiting factor as the flow through the openings is increased. 

Recent investigations 1 * 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 V 2 for high to moderate velocities, with F 1 ' 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 VENTILATION 4 

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). 

6 Carbon 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 t . 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 t = 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. 

C a = 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 / 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 + / 



and for a compound wall of several materials having thicknesses in inches 
of rci, # a , x 3 , 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 



/ 



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 (C a ) for the unit construction, in which case 

-r- is replaced by -~-. 

As in the case of the simple wall, /i and / 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/ are the same, 
and/i = / ; but if the outside air is in motion, then/ 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 Society 1 . 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 / 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 C a , 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 
C a 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 (C a ) OF BUILDING 
MATERIALS AND INSULATORS^ 

7Vr 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 (C a ) 


CIS 8 

1 s ! 

3 3 


AUTHORITY 


MASONRY MATERIALS 


Common 






5.00* 


0.20 




- 


Face 






9.20* 


0.11 




BRICKWORK. 


Damp or wet._ 






5.00 fr 


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 mixes d 

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 


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 




0.66f 


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? i n 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: 

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

2 A. 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. 

6 K 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 ASJB[.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 (C a ) 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 (C a ) 


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 (C a ) 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 




u u _.... 


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 
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 (C a ) 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 (C a ) 


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 


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 (C a ) 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 & ) 


I s ! 

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 


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- 


n7 moisture 


22.0 


75 


0.67 


1.49 


4 




Qcr 


32 


75 


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 


^ 

^^. 


^ 







& 


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 Veneer d 



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 



fl Computed 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 


S 1 




"^ ; 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. 

6 These 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, 3i in. 



104 



CHAPTER 5 HEAT TRANSMISSION COEFFICIENTS AND TABLES 



INTERIOR FINISH 



No INSULATION BETWEEN STUDDING 





e 
o 


i 








c 


2 


^i 


c: 


35 


bfl 


5 


. 


"5 


"3 


^x 


.2 


T!i 




^5 


? 


~ 


& 


. 


S 


^ 


.s 


1^ 


^c 


e^ 


1 





i, 


1 


1 





'.^S 


^ 


|=3 


l|| 





^ 








-S 


3-1 


fo 
S 


S - 


a l-5 


c 


o 


P? 


S # 


5.2 





rt3 


g 


O a 





i 


2 


31 


31 


2-5 


g 


-0_g 




e rs 


.?^| 


jll 





ts 5 


S| 


si 


^4 


i! 


!i 




sr| 


?l|f 






t.2 


2.5 


l'"- 


JSB 


-Is 


|g| 


SJS g 


^T| |^ 


1 


5 


^* 


sS 


5 


5 


I.S 


sil 


slS 


liS-a 


A 


B 


C 


D 


E 


F 


G 


H 


i 


j 


0.25 


0.26 


0.25 


0.19 


0.15 


0.11 


0.19 


0.19 


0.061 


0.17 


0.23 


0.24 


0.23 


0.18 


0.14 


0.11 


0.18 


0.18 


0.060 


0.17 


0.31 


0.33 


0.31 


0.22 


0.17 


0.13 


0.23 


0.23 


0.064 


0.20 


0.25 


0.26 


0.25 


0.19 


0.15 


0.11 


0.19 


0.19 


0.061 


0.17 


0.19 


0.20 


0.19 


0.15 


0.12 


0.10 


0.16 


0.16 


0.057 


0.14 


0.24 


0.25 


0.24 


0.19 


0.15 


0.11 


0.19 


0.19 


0.061 


0.17 


0.30 


0.31 


0.30 


0.22 


0.16 


0.12 


0.22 


0.22 


, 0.064 


0.20 


0.27 


0.29 


0.27 


0.20 


0.16 


0.12 


0.21 


0.21 


0.062 


0.19 


0.40 


0.43 


0.40 


0.26 


0.19 


0.14 


0.28 


0.28 


0.067 


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 


0.066 


0.22 



INSULATION BETWEEN STUDDING 



<f Y"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 towering 1 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 


; 


i ; 


i 








"S 


el to 
















c 




rt 




.?JlgJI 





CO 


M 


W 


S3 


co 

r-t 


T-* 


S 


O 





IH 





J 




^^J *^J 




o 











O 


O 


o 


o 





O 


o 


"5 




pqt-5 fe 


























*o 






























w 


rf 




1 .1 a 




















o 




^* 


*53 
d 


32 




s .l?l| 


fi 


CO 


iS 


a; 


S! 


cc 


i3 


S 


C5 


g 


^ 





1 


e ** 




^ o^f^ j. 




o 


o 


o 














c 




O 


o 


3 


CO 





JS^"s 


























o 


2 


g 


CG 


























g 


H i 





.S JS 


























1 


o 


,_q 


PH bo-a-^'j'D 




fs. 


^, 


o 


Q 


CO 


CO 


** 


o 





o 


00 


rt 


^ * 
. Q 


f=^i 


&.S"MJ3 -S ^| 


u 


(M 


w 


CM 


N 




T-( 


1-1 




o 




o 


.2 


5 


PM 

O 


=|S|gg 




O 





O 


O 


o 





o 


o 











c 
o 


jj 
































2 I 


B 




























JS 


i 


I: 


gg 


























.5 


i 




|-E'i 








w 


c 


3 


IN 


o 





w 


i 


<N 





0) 


2 ^ 











o 














o 





c 


O 


o 


S 


2 1,2 




tS^ 


























& 
































V 


g li 




.2 


























& 












OS 










o 






CO 


CN 




2 So 




^ 


<5 






CO 


CO 


CO 


N . 


<N 


"3 


o 






^ 


O vJ b, 










d 


d 


o 


d 


d 


d 


O 


o" 


d 


d 




,^i fe 2 




o 
& 


























to 


w "S-S 






























*B 


S -"S"** 




























aj 


< ^ 


d 




^ 


N 


W5 


Tt* 


ID 


sO 


t- 


00 


0^ 





^ 


_3 


W ?1 ^ 


12; 
























*N 


"i 


fa "1^ 




























aj 


o ' 


03 


























H 


o- ^ 


O2 














< _ s 












ti 


S?^ 

















_c 


^ 










c 


S: *!?'S 


^ 














^ 


. 


, 


^ 








o 


2 -i-ps 


H 














' 


^ 


IS 


-S 






<TS 


w "o *? 


W 














fl 


^ 


fe 


\> 






^3 


5 *^o 


t> 

















* 


g 


CC 






'i 


S |-g 

















"3 





P 
C 


s 






s 


t?Z fc * 

















SB 


M 


| 


s 

*0 






i 

03 


8*1 

















S 


1 


53 

4^ 


1 






v- a-- 
"3 5 o 


w "J 


DQ 

g 




c 

1 




1 





1 




1 


jD 

1 

S 


1 

S 


x: 



ffl 


1 


V 

1 


I 


ow^r 

^M e 

W *"-C c 
w *o H S 

3 S 3 g 






























rtv *: 


w ^ 












^^ 


.S 














^~" bo fi *5 






























C C i j^ 


^ 












"^ 


Ntt 










'"T" 




= -C^ x 


w 
























S 




* >> M 














O 


1 










S 


.S 


i ^S 















S 














^ 


TS d d 81 5 


TABLE 8. 

Coefficients are e: 




i 

5 
u 


i 


-00 


Lath and Plaster ( in. 


Lath and Plaster 


i 
1 

i 

h 


[nsulation ( } A in.) and PI 


Lath and Plaster 


Lath and Plaster 


Lath and Plaster 


Lath and Plaster 


V 
OB 



e 



r-l 

"2 



oard (2 in.) and Plaster i 


mputed from factors mark 
ickness assumed to be *56 i 
ckness assumed to be % i 
sed on one air space with 
ed from lath and plaster cei 
space faced on one side w 








O 


1 


I 


I 


I 

s 


Metal 


1- 


8 

i 


1 


^ 


.0 

1 


y?l|? 


107 



AMERICAN SOCIETY of HEATING and VENTILATING ENGINEERS GUIDE, 1935 



I! 



u 



O *5 

O 8 



D ft'z 

I 



SS 



M ^ 

U a 

li 



3 'I 





sfiri 




THiHQOCO 


^*coco 




s^ss 


-000 


^ CO CO CO 






^1-So^ 




OOOO 


oooo 


oooo 


oooo 


oooo 


oooo 






CQ"H^ W G 

o 




































S 




^"o 5 
















1 




SHI 


Q 


OiO^J 


S^^ 


CO CO CO CO 


CO CO COIN 


^liM<MO< 


^THrH 


J5 




P||| 




oooo 


oooo 


oooo 


oooo 


oooo 


OOOO 


"S 


1 


" 
















1 




















o 




# a g DC .S 


















S 






^ooo t- 






<N^jHO 


.r^coo 




S 


PSI 

O 







oooo 


oooo 


oooo 


oooo 


oooo 


oooo 


c 

8 




|>cg H 
















1 


E-i 


to 
















> 




cj fe^l, 
















o 






PQ 


Ob- 1C CO 

^cococo 


OOiOCOO? 


ooo^ 


%%%% 


O O5 OOOO 
C<J i-HiH r- 


cococqeq 


^>> 








oooo 


oooo 


oooo 


oooo 


oooo 


oooo 


i 1 




IJS 
















Is 




fl- 
ip f 




IOO5COO) 


C5^O>O 

1C U5 O Tf 


CO CO CO CO 


CO COCO CO 


^sss 


lOT*^ 


^1 

a m 








oooo 


oooo 


oooo 


oooo 


oooo 


oooo 


1 1 




=B ra 
















o o 




O 
















.'f -2 


















M 


1 ^ l 




g a~ 

H &< cQ 
















;.| | 




5 fc K B 




*oooo 


^fOCXDO 


M<OOOO 


^COOOO 


^CDOOO 


TjHCOOOO 


>> > 




g O g 








iH 


f-H 






o o 




















2 " c 




fi O^ 1 
















S s 




















c3 .' rt C 














"O 


1 




*j ^fl *j 










0> 


CJ 


Irt 
















2 


u 

09 





3 


5 ' " 


SjB^jS^ 




~ 






05 


03 


CQ 


CO 


4J 


/^ ti^'O "-< ^ c 


















<U : 


H > cu j S*^ 




5 






fll 


P"< 


o5 


HN 


03 




Jjr 


S 






. "0' 


o 




o 


a) 


. "O 5^5 w 





o r 








st 


09 

OS 


i 


CO 


lass's 



o 




o 

Jr 




o 


ja 


S 


5 


.-< ^ 


Qfi M3 


D 














M 




. T3 , G *C ^ 


' 


V tv ;A / V\A ^i 


J 




J^ 


kJ 


o 


o 


*-*T3 


l&'g&gS 


g. 


/ n*V \\4 


S 
o 




o 


3 


52 


5? 


Is ^ 


|^ S 51 


TYPICAL CO 


fM\ 

ra 

Ww 


TYPE OF 




:er Applied Dir 
ete 


1 or Furred Mei 
letting 


li 
II 


or Furred Cell 
nd Plaster (*4 


in.) on Gorkbo 
Mortar (H in.) 


ed from factors i 
res in COLUMN / 
ss of yellow pine 
res in COLUMN 1 
ss of maple or oa 
ss of tile or terra 








I 


1| 


1- 


1} 


1* 


V 


~> 3 J Z2 o <U 
3 boc bag c 








. 1 


S. 


CO- 


Is 

. g 1 *^ 1 

cc 


03 


IJ 


Vtfrfrfrt 



108 



CHAPTER 5 HEAT TRANSMISSION COEFFICIENTS AND TABLES 










t 


JS 










i 


<j 


o 




*>> 










1 

C3 


& 




d'* g""^ 




o^oo 


**C5 


O OO r-* r-l 


O 




"&3-Q S 5 


K 


OiCiOif 


^00 


I 1 1 ( F-4 T ( 





I i 




*i3sa 




OOOO 


o'o" 


OOOO 





, J *"* 




o 










8 


*"""* 5 














Si 


B 7* 

23 ^ 














c 
















'u 


z jjj 














0> 


I - 




HI 


ft 


00 TH ^fCD 

ooot-to 


coo? 

-COU5 


(M O N W 


11 


w $. 




^^? 




OOOO 


o'o" 


OOOO 


.5 


w * 




o 










II 


b4 t3 

H ! 


O 












s5 


& 


O 












"3,5 

Q 


| 


s 


^cBfll 










".S 


c4 s 


fe 


cs o .- o.g 










'g"S 




'H 














> I 


OQ 


g a g: ^-.S o 


Q 


C^NN W 


^cS 


S iS o O 


'73 "0 


S 5 


2 


"g^S oljl 




o' 0*0*0" 


o'o" 


000*0* 


|1 


-< %> 

*"* "5 


O 


^^f 










. s 
31 


Q fe 4J 


{2 










c * 


D go 


S 


i 








.2 


-I 















|| 


o !l 




JT -2 








05 CO 


IP 


2 ** 












O O O C5 


MC 


2 gc- 




SJ..S2Q.5 2 


tt 


CO CO CO 00 


(NCN 


-* rM 




o Si 




JS S'o'lo 




ooo'd 


o'o" 


o p" o" o" 


Is 


C/3 *tej |g 




2st>^* 










c2 


O*^ "tt o 




l ^" 










^ 


-* R 














. c 


'S'S 














g 


fl 

*-* 














S 6 

.is o 


w <*t^k 




'S' 










T3 


S 5^ 




c 




'-i 






1& 


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* 

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 ^ 












jsJ --3-2 


I 


I r ysfr g z | 


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 Insulation 6 
Rigid Insulation 6 
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' t j/Z_'^&>- 


H 




H . 


c 
o 


c 




.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/^ Uf iTi rj> ri'- 

CtlUKfi'^ 








irVULAUON/ ^ B1 I r ' /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 , l i'i ULAri( '"/ 
lOOHNffj / ROOFlflffl / 


Wood 
Wood 


1* 

1 LjTfc 


5 
^ 


557J ) )T J V 7 j >V /J Si ry / yy yy f y . / .* 

',>' p|-' g ( 


Wood 
Wood 


#* 

4> 


7 
8 


CLtlllHtf^. 








iNJULtftOtt/ Itl/ULAtlfln/ 
RflflRHCi 7 T.OPIM<^ ^ 


Gypsum Fiber Concrete 6 
(2 in ) on Plaster Board 






f f r/y n- :-. - : -V- j ^^ "^ * "" ?: * : '"'' > ] 


(H in.) 


2H 


9 


fLAJTCR. 50AfcP^ PUA/TCR. MAH.P* 


Gypsum Fiber Concrete 6 
(3 in.) on Plaster Board 






CttUWtf'^ ' 


(H in.) 


3% 


10 


MOH r uTi 7 ^g^""( 


Flat Metal Roofs 

Coefficient of transmis- 
sion of bare corrugated 






<^]jjf'^ |'E""?iP r ; 


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 

'.5 s 

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 





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 plaster t 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 fc 



1 

2 "5>lJ 

2 ll 

S ^'S 



l 
I! 



<J 
E 



aJ 



s 



< 



(m 
pnB (Tit z) 



(in 
(mi) uoi 



j pus 
(m f) uoi^nsaj piSrjj 



is Vl PA1 



(m 
(-ut %) 



(pasodxg BJ 
Sxnj] 



O N 



S 

i 



H 



s ill 




2 
"8 



-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/ = 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. Factors 2 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 testing 3 . 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 ;/ = 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 C a ) , 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) ............................................. Ijf 6 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 Z7 C e ( . 

U = ttX/r+Z7ce (6) 

where 

Z7 r = 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 U r and Z7 C e 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; / = 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; / = 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; / = 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 I 1 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_{gL: 


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 plaster b 


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. 36 r 1930). 

120 



CHAPTER 6 AIR LEAKAGE 




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 Pf9 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 Hour 3 - 



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 calked b 


3.3 


8.2 


14.0 


20.2 


27.2 


34.6 




Around frame in masonry wall 
calked b 


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. clearance . In- 
cludes wood frame leakage d 


6.6 


21.4 


39.3 


59.3 


80.0 


103.7 


Ditto, weatherstripped d 


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. clearance 6 . 
Includes wood frame leakage d . 


26.9 


69.0 


110.5 


153.9 


199.2 


249.4 


Ditto, weatherstripped d 


5.9 


18.9 


34.1 


51.4 


70.5 


91.5 




Double-Hung 
Metal 
W T indows f 


Non-weatherstripped, locked 
Non-weatherstripped r 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 
Windows k 


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 
1 Q7 

<u 

02 

< 




















































50.03 
45.63 
40.83 
35.4Q 
ZdSO 










, 
J 






C 


f 


d 


J A 


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 



































































































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 


l Ato % 


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 levels 3 . 

Sealing of Vertical Openings 4 

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- Wells 4 

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 



3 Wind 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 - t ) (3) 

where 

Hi = Btu per hour required for heating air leaking Into building from 
outside temperature t 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. 

t 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- t ) (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 

M e = Vl5 2 + 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 = Vl5 2 - 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 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 SPECiFiED a 



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 / 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 method 2 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 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 SURFACES 3 

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. 


\ 


g o 


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 Concrfca 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-i n - 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-i n 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 

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 


W T 


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 


W T 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: 

H t = 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 Pittsburgh 1 . 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- 
tory 2 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 radiation 1 ' 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 more 4 . 
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 papers 3 ' 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 experiments 4 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 Walker 6 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 Laboratory 6 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 Laboratory 7 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 tests 4 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 = Qd (0 - 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. 
d = 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. H t = 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 d 
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-passed 1 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. 
H s = total sensible heat gain, Btu per hour. 
/ = room temperature, degrees Fahrenheit. 
t y = 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 + = 13.83 

x = 13 - 8 3 * 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. 

H s = H t - 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 

D l = 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 W T hat 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 

165 



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 system 1 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 Book 2 , 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 SiO 2 . 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 

171 



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 

173 



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 

174 



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, 

179 



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 a PP rox * matel y 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 w r asher 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 2 t 931 

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 

T7 1 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... 

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: 

185 



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. 

187 



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 

191 



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 

193 



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 dew r 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. 

195 



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 .. 100 C < 

For two banks, both up-stream - -. 95 r ^ 

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 p 00 m& 




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. 

199 



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. 

200 



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. 

201 



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. 

203 



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. 

205 



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 

206 



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 

207 



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 

209 



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 



2117 



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 RATINGS 1 

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 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 



en 
D 

CO 




8 


*H co o oo t~^ CN 

t~- rH NO CO ON to 

vO Jt s ^ ^^ OO OO ON 


CNOvt^-ONCOvOCNCOrf 

OrHrHCNCNCOCOCO-^l 








O O O O O O 




1 




8 


O\ CN ON CO O * 
co oo CN O Jr- CN 
t^ t^. oO ON ON O 


tO^CNTfOOCNOOrHfN 


CO 
en 










D 
O 

< 




b 


OO OO O O\ O O 


g^sSIS 


> 






O O O -< rH 




O fc, 
to 

s 

td *i?> 




fc 


00 CN f*- tO rH Jf- 
OO ON ON O "rH rH 


vO ON vo CN J> CN -tr rH co 


S S 










w ** 

1 ^ 




s 


VO O rH Tj< rH 

iO O to co ON to 


Ill^lslgf 


^< 1 


1 








s .^ 


1 


fe 


-^ OO f*- rH IO CO 
O O rH CN CN CO 


ssaiis^gi 


III 


fc 

O 








!! 


1 


3 


33885 


i 1 1 g g 1 1 g 


Is ^ 


| 








ferS 

o g ^ 
^ ^i 




fe 


rH J>~ ON tO rH CO 

ON CO 00 I -* O 
rH CN CN CO TP * 


giSSSIsis 


8 "s- 










^sl 




fe 


CO O O O to ^ 


SSigilii 


H H 

tn 










^ 




o o 


ON to vo to TJ< to 

\O Q if) <& -r-4 X*~ 
CO *? ^< to vO vO 


lOCNVOt l>-tOCNO\^ 

!>-VOCOC>vtOOOrHlOO 
<* OO O\ Ov O C^ rH rH CN 












w cq 
Q ^ 







| | g | | 


rHONtO^t^-COrHOOt 1 - 
t^tOCOC>NtOOOrHlOO 
OOOOOrHrHCNCNCO 


1 






rH rH rH rH rH rH 


rHrHCNCNCNCNCNCNCN 


I 




fe 

| 


OO to Q O ON rH 
to to vo *-- ** OO 


IlSssisil 


to 










T-H 

I 


8mm PRBSSOHB 


LB PBR SQ IN. 


O CN IO O to O 


O ^D O C5 O to O O O 



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) 


1 


CO CO TH 


00 to 

oo oo 


1 


o 




s 


to rt* 

"sf ON 


CO 
CM 


CO 

NO 
CM 


CO TH 

00 CM 
CM CO 


ON s 


000 






















s 


TH CM 00 
to ON CO 

f- f-. OO 


i i 


8 


CO 



3 


S S 

CM CM 



CO 


co 

CO 


co co 


"> "a 

S S 


O O O 





















g 


QO OO ON 


co 00 
ON O 



r^ 




2 


CM 


oo to 

*>- CM 
CM CO 


t-~ 

co 


CM 

S 


CM r- 


ftj 

CM fe 



000 




















s 


CM O to 
ON co f- 1 

OO ON ON 


CM OS 


CO 
CO 

, < 


CM 
CM 


S 

CM 


55 S 

co co 


i 


5 


^H CO 

ON CM 

^i to 


S 1 

to o 




\ s 


i 1 i 


CM 00 


1 


CO 
CM 


S 

CO 


NO CO 


to 


1 


NO ON 

to to 


e I 






















W 


i 1 3 


CM Os 


00 


CM 

to 

CO 


i 


oo co 
^ to 


CM 

CO 

to 


i 


ON 1-1 
CM NO 

NO NO 


i 1 




















*"" 8 


. 


3 S 


CO O 
CM CO 


s 

CO 


CO 


S 


to o 
to NO 


1 1 


00 


CO CM 

ON CO 

NO t 


"E 

o 2 
*-. 3 


3 


1 h 


00 to O 

f>- T-H NO 

1-1 CM CM 


a S3 

CO CO 


CM 


t-^ 

$ 





oo co 

CM fi- 
ve NO 


CM 
CM 





g i 


i 1 




















bo 
c 


o 


CO O * 
to ON co 
CM CM CO 


CO OO 
O to 


00 


CM 

to 


t^. 

s 


8OO 
^ 
f- f- 


ON 


CM 
CM 

00 


CO OO 


S 

2 S 

ON 13 


2 


11 


ON CO O 
CM NO -H 
CO co "^ 


fCO 
CO 
to 


CM 
J>- 


1 




TH 


g 

f- 00 


ON 

NO 

CO 


ON 
CO 


i-l to 

ON ON 


^ S 




















, | 


- 


IO CM to 


00 O 

to NO 


1 


to 

CM 


00 


to ON 

00 OO 


CO 

o! 



ON 


CO 00 
CO CM 
ON O 


C 
15 




















CM 

(Tt 


c o 

7 


i 1 g 


t-- OO 
CO OO 
NO VD 


CO 
CM 


1 


s 

oo 


f-, co 

CM f- 
O\ ON 


00 

T-H 




i 


^ CM 

NO O 


a 

S 1 














CM 


CM 


CM CM 


CM 


STEAM PBESBTJBB 
LB PHB SQ IN. 


O CM tO 


O to 

1~( -p-l 


8 


O 
CO 


i 


IO NO 


O 


to 


oo ON 


T3 

8 | 



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 rating 1 . 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 VENTILATORS 2 

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 Ventilators 3 . 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, 

H t = 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. 

H v = heat required to warm air for ventilation, Btu per hour. 
Ht total heat requirements for both heating and ventilation, Btu per hour 

= H + H v . 

Q = volume of air handled by the ventilating equipment; cubic feet per hour. 
t temperature to be maintained in the room. 
t 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; t Q = 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 = 0.24W &-*>)+ O- 24 w i (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 d 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 - 



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 (t y - t ) (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 t Q . 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 


f r"^"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, 

241 



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 
room f 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. 48 1 , as 
follows: 

1. Optimum comfort is the most tangible criterion for determining the air conditions 
within a residence. 

2. An effective temperature of 65 deg 2 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 

1 See 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 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 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 
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 
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 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 

255 



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. 

257 



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. 

259 



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 Baltimore 1 
by actinic methods show that the ultra-violet light in the country was 
50 per cent greater than in the city. In New York City 2 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 Hamilton 3 , Ro'senau 4 , and Henderson and Haggard 5 . 

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 

1 Effects 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 found 6 that 2 mg per day is the smallest dose, by inhalation, 
which in the course of years may result in led 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 groups 7 . 

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 
1 A 


y* 


Phosgene. 


Over 1 A 


1 A 




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 Catchers 8 

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. 

269 



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 
Devices Used in General Ventilating Work 1 explains how such devices are 
rated by (1) capacity in cubic feet of air handled per minute, (2) resistance 



01 AM. 
OF 
PAR- 
TICLES 
IN 
MIC ROMS 


SCALE OF 
ATMOSPHERIC 
IMPURITIES 


RATE OF 
SETTLIK& 
IN FP.M. 
FOR 
SPHERES 
OF 
DENSITY 1 
AT 70 F. 


NUMBER 
OF PAR- 
TICLESW 
ONE CUJT 


SURFACE 
AREA IN 
SQUARE 
INCHES 


LAWS OF SETTLING 
IN RELATION TO 
PARTICLE SIZE 

(LIMES OF DEMARCATION APPOt.) ; 


AIR CONTAINING 
.0006 GRAINS OF 
IMPURITIES PER 


8000 

eooo 


Lu _!__ ^ [_j 





1750 








PARTICLES FALL WITH ! 
INCREASING VELOCITY 


4000 

^ooo 

1000 
800 
600 

400 
200 

10O 
80 
60 

40 

eo 

10 




i 

! ' 






790 


.075 


.000365 


| PARTICLES SETTLE WITH CONSTANT VELOCITY 


C = 24.9V Ds, 








S 


C-Vlocity ft/mm. 

dDiam of par- i 
tide m cm 

D = Diatn of par- 
tick m Microns 

r Radius of par- 
ticle in cm 

g-981 cm./Sfcc 4 
acceleration 

s,-Dtnsit_y of 
particle 

S z = Density Of Air 
(Very Small 
relative to S,) 

air m poises 
l8l4xKT T for 
o.r at 70 F. 

X- I0" 5 cm. 
(Mean free 
path of gas 
molccule.5 ) 


1 15 


-TT- | 


1 




^ 


~ 5e ' g 


V 




555 


.6 


.00073 














_a 




L_" =. 

64 ^ 


Qr-< 

li 

5-1 




59.Z 


75 


.00365 S 




-| 
o: 
q 


UJ 

X 




1 " I 


c 
O w 


STOKES 
LAW 

c ,gr*gs.-3 t 

FOR AtR AT 70* F. 
C -300,4605^* 
















14.S 


600 


.0073 




















p/ 




w w "j w 


U 


x 


., 


75000 


.0365 S 






^ 




J2 r- 3 ^ 


V) 


U- 






5 




5(5 " 


^ 


V) 

nJ 


V) 


O 

u. 


1 






l | 2 


E 


a 


6 

4 

Z 

.8 
.6 

.4 

.e 

j 












j ij 






J48 


-00,000 


.07-3 








J2 


ua u 






.007 5" 




.365 S 








^ 


5 u 










tf 


a. 


3? s S 


a 











a. 






u 




















.OOZ=U4" 


60X10 7 


JIMS*. 
.73 


CUNNINGHAM'S 

FACTOR 

C'-C OF STOKES LAW 
K-.8T0.86 


















vi 




X 




i 


I, 






PER HR. 


75X10* 


3.65 ^ 


r 


(0 


a. 


UJ 


z. 


i 






u. 




UJ 

t 


| 


S 




? Z 










1 

a 


S 


5 











.01 

.001 










b==J 


T F 






O 


60*10* 


7.3 


PART ICLES h 
GAS MOL 

BROW Nt AN 
MOVEMENT 


tOVE LIKE 

:CULES 

A* Distance of 
motion m time t 

R-Gas constant 
* 6.316 x O 7 

T - Absolute 
Temperature 

N Number of Gas 
molecules m 
one mol6L06)ci0 23 




u. 


ij 


K Sy J 


& 







75xtO lZ 


36.53- 




z: 


j 


-5g ^ Si 


u 


UJ 




f= 


od^i g s 


o. 




O 

Z 


1 


S 


S"l^ ft ^ 


o 






2 




"'o H,. _j^ y 









fcOXIO 13 


4 
73.0 




7: 




|zH.:S 


-1 







75X10 1S 


365 ^ 




< 
-S- 
en 


-J 


-1 


"5 uj 5 














-E 








tu 







Compiled by W. G. Frank and Copyrighted. 

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. 



K; X.U 

$ 




\ 


\ 




































i \ 


*-? 


i 7 


us- 

~YP 


r H 

E \ 


OLL 
'I5d 


?;/V(3 
:ou 


c* 

5 F 


\PA 
-fLT 


err 

ZF5 


y o 

IN 


F L 
PC* 


wr 
;nz> 


r 

5 




^,, 




\ 


y 


V 




























r;: 




\ 


) 






























V} "^ 






V 


\ 


\ 


























% 






\ 


v 


V 


k 


> 






















* 








% 


^ 

^ 


^ 


\ 


x 




















* 










\ 


\ 


X. 


\ 


^ 


~~~~*~-. 


-^-^. 


^^ 












-J & 

s 

tft 2 
















^ 






--, 









~-__ 


. 




3; 





































Of Z 3 A- 5 4> 7 a 9 to // /Z /3 14 15 H> 
\A/EKS Or 60 OPERATING HOURS 

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,} 

0.4 200 2 



0.16 7 2 2 

F 2 2 = 16,000 
F 2 - 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 

279 



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 

281 



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 



1 See 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 Blowers 2 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 satisfactory 7 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 


c 
*l 








Arr. 4* 
For direct drive. 
overhung. No bearii 


mounted on motor 
Pedestal for motor 












-" \ 







n\=i 



3 a 





Its 





Similar 
on fan, 
couplinj 




& 

o 



9 
w 

i 

P3 

04 
< 

O 

2* 




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. 

297 



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) 

299 



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 
noise 1 . 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 

301 



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 = I 1 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 + <xS z + a 3 5 3 -f , wJiere Si, 5 2 , 5 8f 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: 

J l = /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 


51 


0.72 


Acousti-Celotex, Single B % 


0.11 


0.45 


0.68 


Acousti-Celotex, Triple B 1% 


0.20 


75 


0.67 


Acoustic Flexfelt 


0.27 


0.56 


0.68 


Acoustone 1 




66 


0.69 


^Vkoustolith plaster 3^ 


0.21 


29 


37 


Akoustolith A, Tile 1 


0.14 


0.48 


0.83 


Brick wall, unpainted 18 


0.024 


031 


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 


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 

7 11 ~ 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 sabines 2 of absorption ; and that the noise level in 
the equipment room is 70 db above zero level. 

rii 
70 - = 10 logio -y- (from Formula 1) 

Tfl 

* = 1Q7 



~ = 10 7 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 (/" = 10 5 ) the noise level due to all of the noise entering through the 
wall would be found as follows: 



-^- = 10 5 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 (/" = IO 7 ) and suppose that there is an assumed 
ttenua tion in the duct equivalent to a transmission factor of 0.0002. Then, 



IO 7 X 0.0002 = 2 X IO 3 



-f- 20 (from above) + 2 X IO 3 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 
Soderberg 3 . 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 

t r = 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 MATERiALS a 

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 10 5 
0.25x 10 5 
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 10 5 


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 10 s .) 

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 10 4 . 

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 10 fi X 1.86 X 10 4 = 0.28 X 10 9 mechanical ohms (or absolute 
units). Therefore, 



V 



<- 28 X 



+ 



X 100 X 0.134 



(0.28 + 10 9 ) 2 4- (2x X 100 X 4.54 X 10 6 - 



= 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 10 5 X 645 - 0.97 X 10 7 mechanical ohms (or absolute units). 
Therefore 



-v 



(0.97 X 10 7 ) 2 -f 10 



X 100 X 0.39 



(0.97 X 10 7 )* + ( 2x X 100 X 4.54 X 10 s - 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 buildings 4 . 

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 experiments 5 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 
wall r 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 3 r 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, I 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. 



10 7 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 J s , /b, and / v the intensities of the street, banking room, and ventilation noises, 
respectively, and J the reference level, we have: 

/o ~7o~ !o 

The total intensity, I, will be 7 S -}- I b -f 7 V 

The intensity level is 10 logio -j- 

=101og 10 ^dl^_/v) 

= 10 Iog 10 (10 4 + 10 6 -f- 10 5 ) 
- 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: 

PS 1 
I = before applying material 

PS 1 
/i f- after applying material 

PS 1 

JL = a = J*!_ 
It PS 1 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 

/ a 1 1008.7 



Ji a 73.4 



13.7 



Noise level reduction = 101og 1P -=- = 10 log 13.7 = 11.4 db. 

315 



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- 

317 



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. 

319 



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 

321 



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? 

323 



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 H v 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 

H v = 5.2 -^ 

9 
therefore, 



V = 1096.5 .J^X__ (1) 

I p 
where 

V velocity in feet per minute. 
h v = 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) 

325 



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 

327 



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 

329 



AMERICAN SOCIETY of HEATING and VENTILATING ENGINEERS GUIDE, 1935 




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, 

331 



AMERICAN SOCIETY of HEATING and VENTILATING ENGINEERS GUIDE, 1935 




e-4 
o 
o 



o'P 






It 



Q 

S5- 
< 

5 

< 
S 



O 

to 







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 



3 t-> eo Os O f* *> <*3 * w "O t~- * 
< M pl CN f*J ro *O f) f*5 fO fO fO fO O 



0.0*00 
r^ f> ^tl * 



O O -< e^i N o ^< 10 o < 



o. o o -* <* 
-< c* ts t> r 






u--O -Ol-* 



O CS C4 C4 



2^,i2 I2-i--^d MSn*N t?J?I 



SSRS 



3 



3*~*.OO oCOOOOOv QsOsOO O -- < 



J T >O OO O O* ^*" 
4CiC^ MOOO 



334 



CHAPTER 20 AIR DUCT DESIGN 



a 




so O * ** 

S S S 8 S 


, 

; 




-* to vo - t~ oo oo 


s 




s ss;s ^&s 






OC O^OH CMCMC"* t^rH 


*"- 




ao acoooc SoxS SSoJ 


s~+ ' 




sj t^^o \ooso -*PO 


"U 1 ^ 

CJ ! *"" 

3 1 




t-~ OOOO0 OOCOOO SO OO OO OO^Qk 


. 


o 


^50X CKO> 0->^<* Ovt- OtO^. 


c - 


<M 


ri^io ot-ec o*O^i r^(*5-* io\o~ 


o 




t~r^ r-r-.t- r-DooO ooeooo oOcooo 


u 


-* fM r5 




T s 


\p t oo Oi 
O vO so 


rs" PO' ^ ^ *^ oo" oo* Ov" 0* -4 tM* P^" 


1 




** f. ^*>.r. r~.r^*^ t^t^-oo oOoooo 


o 




vOsOiO ^rOO< -*000 \OlOfO 00 


s 


ft 333 3315 


SS3 S^^ SJ2S SieS f^^S 


M 


O-Hf* POPOPTl POMCS 




i-3 


ESS 3SS SSS 


*tlOIO SOJt^-OP CNOO * C< *i ^^tO 


1 a 

^ s 1 1 




*OOM *-vOoo Ors-* soooo eN**O 


i T 










00 000 OOOv -.00 01010 


u 3 




Jo s;^^ ^33 ass sss 


' 




vo oseooo oot- ^.^^ PO^HO 


< 




%Z S?SS ^Jg 333 333 









p 3 




338 S3 ^^^ 3^ 3S3 


H 






u S 

S3 ^ 


5 


U3S ^^K S?SS SJSS; 3$S 




O 1-1 


es oo^oo io-^c4 ot * v 


fc. 2 

O 




ij< ^- ^f 'W vo lO 1O*O>O 101010 tovoiO 


tn 


00 01 


OvOOv Ooes Oio*J OOf<- OC^O 




353 


^i^i^t ^i^,^, iotOtO tovotO (O^OtO 


^ 






g, 




333 SSS SSS SSS 3RS 


td 


SS SS5 


53S =555 5SS SSS Sidsi 


j 






u pj 


SSR SSS 


535 333 SS5 9SS SKK 


CJ 1M 






rH 

w 


PO <$ PO PO 5> PO PO 


SS5 335 5^ ^55 SSS 


J 00 

ffl ?J 


o 55??55 3J 


S SS 899 335 999 5S5 


s 


RA8 SSS SSS 


vo*-.oo gj 355 333 355 


jjl 







SS5 



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 3 a 



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 


29 1 A ', 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 in.; 150 ft of 56-in. pipe is 

. , . . 150 X 12 ooo,r a 

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 (1 1 A 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 Blowers 1 , 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 4 a 



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 4 ? 

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 CONSTRUCTION 2 - 



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\TUSTRIAl, 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 conditions 1 . 
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 follows 2 : 

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. 



2 For 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 45 

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 al 3 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 uniform 4 . 

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 - 0.5A V . 

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 
dilution 5 . 

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 system 6 . 
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 yy <#* (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. 



*DaHaValle F 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 \/ h v . 

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 sheet 1 
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) M (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 = d Q . 
Qo = volume of outside air to be introduced, cubic feet per hour. 
d density of air at t , 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; d 0.08276 (Table 3, Chapter 1); 
Mo = 0.08276 X 600,000 = 49,656 Ib; t = 70 F; t = 20 F; H 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 

t y = 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 

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 M r (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. 
M = weight of outside air, pounds. 
M r = weight of recirculated air, pounds. 

Example 8. Using the data of Example 2 and an outside temperature of 20 F t what 
will be the values of M, M and Af r ? 

Solution, d - 0.07495 ;&> =jQ,QS276;() = 110,400; Q - 90,000; H = 100,000. 
_ 100,000 



~ 0.24 X (120 - 70) 
M G = 0.08276 X 90,000 - 7,448 lb 
M T - 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. 
jlf ="" 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. 
t y = temperature of the air leaving the duct outlets,,; 
t K average temperature of air entering tempering coil. 
&# temperature of air entering washer. 
0.24 = specific heat of air at constant pressure 1 . 

366 



CHAPTER 22 FAN SYSTEMS OF HEATING 



Rolling Shutter- 



/ Steam 
Control Valve || _ Control Valve 



-Air Leaving Fan at t y 




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 t x t . 

- *o) M (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 F; tz = 0, t y / 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 t y 
^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) M r 
:..',. 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 t y 
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 M 
- 0.0864 X 90,000 7776 Ib per hour are required, measured at 65 F. 

Mr - M - M - 38,889 - 7776 = 31,113 Ib 



Hi 



7776 X + 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 - t )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- H 3 + H 3 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 t ) 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 t y will ordinarily be different for each room 



With 



se K. (Fig. 7J ine temperature r y 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 - t z }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 hf g 



W = weight of condensate, pounds per hour. 
Q total volume of air, cubic feet per minute. 
AJ = temperature rise of air, degrees Fahrenheit. 
Af g 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 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; t y = 116 F; 
, = 70 F; Q = = 55 l - 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 fr g - 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 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; t y - ^? + 1 = 55 ' 2 Q ^^ 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; hf s 968 Btu, from steam table in Chapter 1. 

w 60 X Q X A t 60 X 5400 X 96.6 , , 

W = 55.2 X fe g = 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 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; M = d X Q = 0.07495 X 60 X 1000 = 4497 lb per hour. In this case 
d and Q are figured at 70 F. H == 0.24 (t - / ) ; M = 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; - - = 55 f - 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; A t = H4 - 39.6 - 74.4 F. H - 60X6912^74.4 
per hour. 



373 



AMERICAN SOCIETY of HEATING and VENTILATING ENGINEERS GUIDE, 1935 

11 When the outdoor wet- and dry-bulb temperatures are 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 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 = H = 0.24 (t - t ] M Q . By Equation 6, M = 
d Q = 0.08636 (from Table 5, Chapter 1) X 10,000 = 863.6 Ib per hour; t = 75 F; 
to = 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 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 

375 



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