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MECHANICAL EQUIPMENT 
OF BUILDINGS 



HBITIKO AlID VEHTILATIOH 



POWBK lOAKTS AND KBFRIGBIUTIOII 



VoL tH—In PrrparaUm 
HISCBLLARBOHS BOlLDlnO BQUIPHBnT 



MECHANICAL EQUIPMENT 

OF BUILDINGS 



A REFERENCE BOOK FOR ENGINEERS 

AND ARCHITECTS 



BY 

LOUIS ALLEN HARDING, B.S., M.E. 

Member American Society Mechanical Engineen; Member American Society Heating and Ventilating Engineera; 
Fonn«1y Piofewor of Mechanical Engineering. Penna. State College; Profeasor of Experi- 
mental Mechanical Engineering. Univenity of Illinois; Chio Engineer and 
Member Firm John W. Cowper Ca, Buffalo. N. Y. 

AND 

ARTHUR CUTTS WILLARD, S.B. 

Member American Society Heating and Ventilating Engineers; Formeriy Aflistant Professor of Mechanical 

Engineering. Geoige Washington University, and Sanitary and Heating Engineer. U. S. 

Wkr Department; Assistant ProfessOTof Heating and 

Ventilation, University of Illinois 



VOLUME IJ 

POWER PLANTS AND REfRlGKRAtlON 






* * 



FIRST EDITION 
FIRST THOUSAND 



NEW YORK 

JOHN WILEY & SONS, Inc. 

London: CHAPMAN & HALL, Limited 

1917 






Copyright, 1*9 1 7» by 
LOUIS ALLEN HARDING 

AND 

ARTHUR CUTTS WILLARD 



• ; . • • • . . 









• • 



* 



PI^BLISHERS PRINTING COMPANY. NEW YORK 



y 



PREFACE TO VOLUME U 

Tms book is a new d^Muiure in tJie literature on the mechanical equipment of buildings. 
It piopoflea to deal not cmly with power plants and refrigeration, to which this volume is devoted, 
but abo wiUi the heating and ventilaticm of buildings which is considered in Volume I already 
publiflhed. In addition to these two volumes a third volume on elevators, lighting systems, 
sprinkler systems, vacuum cleaning and plumbing is now in preparation. 

In order to make Volume II complete in itself it has been found advisable to reprint (with 
minor changes) the Chapters on Heat; Water, Steam and Air; and Fuels and Combustion, and 
parts of two other chapters, aU taken from Volume I. This means that about 100 pages have 
been added to this v(^ume in <Mrder to make it a complete individual unit avoiding the neces- 
sity of constant reference to Volume I. 

The object of the authors is to produce a reference boc^ for engineers, which will contain 
sufficient theoretical and commercial data for practical use in the designing room, and at the 
same time serve to show the student of this subject the relation between the theoretical principles 
involved and their practical application to actual problems. 

All available sources of information relating to this field of engineering have been drawn 
upon, and credit given in the text, wherever such information is introduced. The authors have 
found it necessary, in their own experience, to make extensive use of manilfacturers' data in 
designing the various mechanical systems or plants required in modem buildings. They have 
there f ore not hesitated to include such data in the text in order to illustrate and facilitate the 
design of similar systems in each subject treated. 

References to specific makes of such equipment have not been intended as in any sense ex- 
clusive of other equipment of the same sort, but merely as indicating that the equipment named 
and described is as satisfactory as any to be obtained in the market. 

The authcMn are especially indebted to Prof. G. A . Goodenough for many valuable suggestions, 
as weO as permission to make use of his latest tables of the properties of steam an 1 ammonia 
and also of air and vapor mixtures. 

The Authors. 
Ubbana, III., 
April, 1917. 



... 

lU 



TABLE OF CONTENTS 

VOLUME II 

POWER PLANTS AND REFRIGERATION 

CHAPTER I 

PAGE 

Phtsicai. Units and the Measurement of Heat ...... 3-15 

CHAPTER II 
Water, Steam and Air 16-61 

CHAPTER III 
Fuels ani> Combustion 52-73 

CHAPTER IV 
Bori.KR8 AND Rules for CoNSTRumoN 74-141 

CHAPTER V 
Mechanical Stokers 142-147 

CHAPTER VI 
Superheaters and Economizers 148-158 

CHAPTER VII 
Chimneys for Power Boilers 169-195 

CHAPTER VIII 
Mechanical Draft 196-207 

CHAPTER IX 
Feed- Water Heaters and Feed-Water Purification 208-228 

CHAPTER X 
Steam Engines 229-294 

CHAPTER XI 
Steam Turbines 295-338 

V 



VI CONTENTS 

CHAPTER XU 

PAOS 

Pumps 399-386 

CHAPTER XIII 
Steam Condenhers 386-440 

CHAPTER XIV 
Cooling Ponds and Towebs 441-464 

CHAPTER XV 
Pipe, Fittings, Valves, Coverings and Accessories 465-524 

CHAPTER XVI 
Power-Plant Piping 525-559 

CHAPTER XVII 
Arrangement op Steam Power Pij^nts 560-579 

CHAPTER XVIII 
Coal and Ash Handling Machinery 580--614 

CHAPTER XIX 
Isolated Power-Plant Data 615-629 

CHAPTER XX 
Cost op Steam and Gas Power Equipment 630-638 

CHAPTER XXI 
Units Employed in Refrigeration Practice 639-642 

CHAPTER XXII 
Heat Transmission and Constritction of Cold Storage Walls . . 643-664 

CHAPTER XXIII 
Heat Transmission of Piping as Used in Refrigeration Practice . 665-669 

CHAPTER XXIV 
Methods of Producing Artificial Refrigeration 670-671 



i 



CONTENTS vii 

CHAPTER XXV 

PAGE 

Cold Air Machinkb 672-674 

CHAPTER XXVI 
CoMPRGfiHiON Machines . . v 075-694 

CHAPTER XXVII 
V^ACuuM Machines 605-699 

CHAPTER XXVIII 
A30I0NIA Condensers 700-705 

CHAPTER XXIX 
Brink Cibcui^ting System 706-712 

CHAPTER XXX 
The Ammonia Absorption Machine 713-728 

CHAPTER XXXI 
ICE'MANWAOTUaiNQ PLANTS 729-747 



y 



Volume II 

POWER PLANTS AND REFRIGERATION 



Mechanical Equipment of Buildings 

VOLUME n 

POWER PLANTS AND REFRIGERATION 

CHAPTER I. 

PHYSICAL XmrrS and the MBiCSUREMBUr OF HBAT 

FUNDAMENTAL UNITS 

MoDBRN engmeering practice depends on the correct application of basic principles already 
devdoped and established in such branches of physical science as mechanics, thermodynamics, 
hydraulics, physics and chemistry. As reference will be made to these fundamental principles 
from time to time, it is necessary to define the imits in which the various quantities dealt with 
will be measured. 

In this country the system of imits in general use by engineers is known as the FooirPound- 
Second SytUm, and the following definitions and examples will show the significance of each 
unit. A table of equival^its (Table 2) is also given, so that the value of the more general com- 
pound units can be foimd in terms of various other imits. 

Defiiiitioiis of Units and Terms Employed in the F. P. S. System. The unit of time is the 

eecood, which is equal to ^^^ part of the mean solar day. i =» time. Time is also expressed 

ciO,4UU ^ 

in minutes and hours. 

L » UngUi. The unit of length is the foot » 0.3048 meter. 

W s weight. The unit of weight is the pound *= 0.4536 Jcilogram. 

A « carea. The imit of area b the square foot. The unit often used is the square inch. 

V = vclume. The imit of volume b the cubic foot. Volume equals area X length = il X L. 
In calculations involving quantity of air required Q is often used for cu. ft. 

Example. The volume displaced per stroke by the plunger of a pump, if the diameter is 6" and 
the stroke is 12", is >i«- X 6* X 12 - 339.29 cubic inches or 0.196 cu. ft. 

If the plunger makes 30 "working** strokes (not revolutions) per minute, then the plunger "dis- 
placement" per minute is 0.196 X 30 » 5.88 cu. ft. One U. S. gallon ■> 231 cu. inches or 0.1336 cu. ft. 

e QQ 

This pomp will therefore theoretically deliver — ~— or 44 gals, per min. The actual delivery of the 

0.1336 

pump will be somewhat less owing to "slip,** which is the leakage back through the pump valves, around 
the plunger, and that due to imperfect filling of the pump cylinder on the suction stroke. 

D=^ density. The weight of a unit volume (l cu. ft.) of a substance is called its "density." 
The density of water at 70^ F. is 62.3 lb. per cu. ft., and at 60^ F. =62.37. 

The densities of the following liquids at 60^ F. are: 
Petroleum: 48.7 to 54.9 lb. per cu. ft 

3 



4 POWER PLANTS AND REFRIGERATION 

Mercuiy: 848.7 lb. per cu. ft. Specific gravity - 848.7/62.37 » 13.6. 

The pump in the preceding example would, therefore, handle 5.88 X 62.3 or 366 lb. of water per 
minute. 

If the water end of the pump were operated by a steam cylinder having a displacement of 0.349 
cu. ft. per stroke and took steam at the same pressure for the full stroke as in the "direct acting" type 
and assuming that the steam pressure were 100 lb. gage, we find from the steam tables that the density 
of steam at this pressure is 0.2566 lb. The "steam consumption" of the pump, therefore, would be 
0.2565 X 0.349 X 30 X 60 or 161.6 lb. per hr. theoretically. 

V = velocUy. The rate of motion of a body is measured by the distance passed over in a 
unit time.' Velocity is expressed in ft. per sec. 

a =s acceleraiian. The rate of change of velocity measured in ft. per sec. is termed accelera- 
tion, and is stated in ft. per sec. per sec. (generally expressed, ft. per sec.^). Acceleration may 
be either positive or negative, depending upon whether the speed of the moving body is increas- 
ing or decreasing. The uniform acceleration due to gravity, denoted by the symbol g, is the rate 
of gain in velocity of a freely falling body and is 32.174 ft. per sec.' The value of ^ is generally 
taken as 32.2. 

M = Tna88. The expression W/g is termed "mass." A unit of mass is the quantity of matter 
in pounds to which the imit of force (1 lb.) will give an acceleration of 1 ft. per sec.^ 

Relation between Velocity, Acceleration, Time, and Space Passed Oven When the ac- 
celerating force is uniform the acceleration will be uniform. The velocity at the end of t seconds, 
if the body starts from rest, will be 

V ^ at; whence a = — , and / = — 

t a 

The space passed over at the end of t seconds is equal to the product of the mean velocity 
and the time, or L = J^ t;<, or L = )^ at^. 

A force of 1 pound when applied to a mass whose weight is 32.17 pounds will produce an 
acceleration of 1 ft. per sec.* when the mass is moving against no resistance (frictionless 
motion). A force of F pounds acting on a mass of one pound will produce an acceleration of 
F X 32.17 ft. per sec.' 

The relation between force, mass, and acceleVation is given by the equation F = Ma = Wa/g. 

Substituting the value ^ a in terms of v we obtain 

„ Wv 
F ^—r 
gi 

V = energy or work. The unit of work is the foot pound, and is the quantity of energy 
expended or the work performed by a force of 1 pound moving through a distance of 1 foot in 
the line of action of the force. 

Power is the rate of doing work. Note that "power" involves the factor "time" and is equal 
to the amount of work done divided by the time required to do this work. 

^.p. = Horsepmoer. The unit of power is the "horsepower" and is the performance of 
work at the rate of 550 ft.-lb. per sec. or 33,000 ft.-lb. per minute. 

Example. Required the theoretical work and horsepower developed by the water end of the 
pump in the preceding example if the head or the height pumped against is 200 ft., assuming no frictionul 
resistance to be overcome. 

The work Um performed per minute is the lifting of the weight of water, W «■ 366 lb. per minute, 
through a height of 200 ft. is 

Um 73,200 

C/w = 366 X 200 = 73,200 ft. lb. per min. and h.p. = — "ii^ = ^^ ,,: - 2.22. 
^ 33,000 33,000 

The actual power required will be somewhat greater, as we have neglected the force required to 
overcome frictional resistance, and the force required to accelerate the water from a state of rest to the 
velocity at which it is delivered. 



PHYSICAL UNITS AND THE MEASUREMENT OF HEAT 5 

The total fonn F requirAd on the plunK^r will be the total preamre produced od the plunger by 
the'water column, ne^eeling friction and the acceleratioK force. 

The prevure per unit arsa produced by the vater column ia equal to the hei^t if in ft. multiphed 

by the density £> of the water orP-flD-200X 62.3 - 12,460 lb. per sq. ft., or p - ' - 86.6 

If A — area of plunder in sq. in., then F — p A 
The work pwformed per stroke of the plunger a 
ft. or i; - 2.446 X 1 - 2,446 ft.-lb. 

The work per min. — I/„ - 3,446 X 30 - 73,380 and tlio powor required ia ^,'' ^, " 2-22 h.p. 

HeaMtremenl of Pr«Miir«. It is cuetotnary to meumire prvsiure by mcaiiB of gages whiofa 
in reality only indicBite the difference between the pressure being meaaured and the preeeure 
of the atmosphere (barometric preaeure) at the same time and 
place. These gages may indicat« either a higher or lower — 

pressure than that of the atmosphere; in the former case 
tbey are known as premure gages and in the latter as vacuum 
or dn^fl gaga. 

Prmnm and Vacuum Gages. The most common type 
of pressure gage (Fig. 1) ia provided with a flexible hollow 
brass tube of oval cross section known as a Bowdon tube. 
When subjected to pressure, this tube tends to straighten out 
and thus causes a sector of a gear to mesh with a small pinion 
oa the same shaft with the indicating hand or pointer and ro- 
tate the latter a corresponding amount. The pointer is pluced 
just in front of a graduated dial (not shown Id the figure) 
fitxn ^lich the pressure may be read in suitable pressure 
units such as pounds per square inch. ^°- ^- ^iNotB Spmno PitiaBuitH 

Th«e gages may also be used for indicating vacuum or ''*''"■ '""■"™ ^'»*- 

tM«sauree less than that of the atmosphere. 

Draft Osge«. The measurement of pressures but slightly above or below the atmospheric 
pressure (barometric pressure) is usually accomplished by the use <rf a draft gage (Fig. 2) con- 
nect«d at the stop cock oo the right hand side. 

This is essentially a U tube, containing either water, kerosene, alcohol or mercury, mount«d 
upoD a paduBt«d scale, and reading either in inches of fluid or in pounds or ounces per squoro 
inch. Since the pressure indicated is a differential one, due to the left hand leg being open to 
the air, the reading must be Uitained by adding the depression in the left hand leg to the ele- 
vation in the right hood leg; using sero as the reference point in both cases. Thus, for the gage 
shown in Fig. 2 the reading is 0.95 + 1.05 - 2.00 inches of mercury from which the vacuum or 
pK^re below the atmosphere in pounds or ounces per square inch may be readily calculated. 

Barometerv. The pressure of the atmosphere is usually measured by a mrreurUil barometer 
(Fig. 3) which in its simplest form consists of a glass tube about 3 feet long, closed at one end, 
which after being filled with mercury is inverted in a shallow bath of mercury. The pressure 
(rf the atmosphere at sea level maintains the mercury column in the tube about 30" above the 
level in the cist«m. The barometric height or length of this column of mercury varies with the 
altitude above or below sea level. 

When the mercury in the tube falls, that in the cistern rises in corresponding proportion, 
and vine vcma, so that there is an ever-varying relation between the level of the mercury in the 
tube and the mercury in the cistern, which affects the accuracy of the readings. It is therefore 
necessary before reading the height of the mercury column on the stem of the barometer (Fig. 4j 
by means of the movable vernier C to adjust the level of the mercury in the cistern. 



8 POWER PLANTS AND REFRIGERATION 

Tbe cist^n (Fig. 5) consiats prinuirily of a heavy walled glaas cylinder AA allowing the mir- 
faoe of the mercury B to be clearly seen. 

This cylinder i» aecurely held between bolsters CC in a. movable frame suitably mounted in 
tbe base of the instrument. 

The bottom of the frame D iffeonneclAd with the threaded stem E which raids in the knurled 




Pia. 2. Draft 



FlO. 3. filHPLI BABouenB. 



nut F, by means of which the cistern is moved vertically thus raising or lowering the mercury 
level and adjusting same to tbe tip of the ivory pointer G, which is the lero of the scale. 

All standard or obaervatory barometers of the mercurial type possess this adjustable feature. 
Barometers of other types, such as the Aneroid barometer, must be frequently compared with a 
standard meruurial barometer in order to check the accuracy of their reading 

Bunmetiic Pressnre. By barometric height is meant the hei^t of a column of pure mer- 
cury at 32° F. which just balancee the pressure of the atmosphere at the time and place of the 
obeervation. The ttandard or normal boromefnc ■preavre is defined as the pressure of a column 
of pure mercury 700 mm. (29.92 inches) high at 32° F. This is the norma] barometric pressure 
at latitude 45° and sea level. Sin(% the weight of 1 cu. in. of mercury under these same conditions 
is 0.491 lb. then the normal barometric pressure equals the height of mercury column X wdght 
per cubic inch, = 29.92 X 0,491, or 14.7 lb. per sq, in. 

This pressure of 14.7 lb. per sq. in. is known as the absolvte presiure of the atmosphere at 
latitude 45° and sea level. Now, since the ordinary pressure gage measures only pressures above 
or below that of the atmosphere it is necessary lo add the barometric pretture at the place in 
queetion to Ihe gage reodtnii to obtain the total abadiiU preefure corresponding to the preeeure indi' 
cated by the gage. That is: absolute pressure = barometric pressure + gage pressure. 

The pressures used must be in the same units, and may be expreeeed in pounds per sq. 
Ft ,, P or specific preaaure, or in pounds per sq. in., p, the usual unite for expressing gage prtatute 
P " 144 p. Also pressure in inches of mercury X 0,491 ■- pressure in pounds per sq. in. 



PHYSICAL UNITS AND THE MEASUEEMENT OF HEAT 7 

HEAT 
Doflnllloii of Haat Htal is a form of energy, and not a eubotance. It ia, in fact. Qui 
kinetic and potenti&l eaerE^ ot the molecules <rf which all subetaaoee, whether wdid, liquid, <a 
gaseous, are compoeed. Whenever tiie vibrattny 
motioii of the moleculea composing a body of given 
masB ia increfksed from any cauae the thermat kinttie 
tntrn ia increaaed. The temperature of the body 
liaea, ita tentible heat increaseB, and the body feeb 



Hie thermal potential energy of a body of given 
mass may be increased by causing it to expand ac 
change its state, thus separating the molecuks 
apinst their mutual attractions and requiring the 
expenditure of woric or its equivalent in heat. The 
work expended in separating the molecules due to 
czpansioa, or in changing their state of aggr^ta- 
tion, as in changing from soUd to liquid, is stored 
in the body as potential energy. There is no 
dtange in temperature during changEfl of state, 
henoe the kinetic energy and the temperature le-- 
main constant. 

Furthermore, the thermal kinetic energy ui a 
body for a given rate of vibration of the moleculea 
will vary with their number or the mass of the 
body. Hence if the rate ot this molecular vibra- 
tioD is the same in two different masses of the 
same substance, they will have the same heat in- 
tetunty or temperature, but the larger mass wiQ 
have the greatO' Aea( cordent or possess more heat 

Heunremant of Tempermtnro. (Thermom- Via. 4. Bran . Via. 6. Canm 

— ■" OOMHTaCPnOM. 



etry.) Inteneily o! heat is measured by (Aer- ' 
momter* and pyromelers, the Utter being used for o™.avAToi« B*ao«-iTB. 

hi^ temperatures, above 400° to SOO" F. In 

engineering work mercuric thermometers are very largely employed. These depend upon 
the uniform expansion of mercury to indicate changes in temperature. Ihe unit <rf measure- 
ment is called a degree, and is capable of very exact determination, provided two points, at 
which (he heat intennty is always constant, can be used as a base or reference for caUbration. 
The meltin^ptHnt of ice and boiling-point of water at atmospheric pressure ore usually selected 
as ba«B, and the uniform expansion of the mercury between these two points is indicated on a 
scale divided into 180, 100, or 80 divisions. (Fig. 6.) Each of these divisions is known as a 
degree and the scalea used are known respectively as Fahrenheit, Centigrade or Celiiut, and 
BAiuniur. The fcTmer is used almost exclusively in engineering work in this country. 

Due to variations, under service. In the glass of which mercurial thermometers are made, 
it is necesary to compare them from time to time with a standard or to check the melting and 
b(MHng-point readings for accuracy. This e^^tration of a thermometer, as the process is called, 
diould always be made before taking any important temperature readings with the instruments. 
The correctioos are either tabulated or plotted, and the sign (+) or (— ) prefixed, the former 
indicating that the correction is to be added, and the latter that it is to be subtracted from the 
observed reading. For emmple, if the thermometer actually reads 200°, and the corrocticai 
taUe draws + 2.3 at this observed temperature, the actual teinperature is 202.3". 



8 



POWER PLANTS AND REFRIGERATION 



A further correction for stem exposure must also be made in very exact work, due to the 
fact that thermometer scales are graduated to read correctly for total inunersion, that is, with 
bulb and stem of the thermometer at the same temperaturei and they should be used in this 



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way when compared with a standard thermometer. If the stem emerges into space either hotter 
or colder than that in which the bulb is placed a "stem correction" must be applied to the ob- 
served temperaturCi and is made by use of the following formula: 

correction = 0.000085 N {t - Q, 

where the decimal is the difference between the coefficient of expansion of the mercury and the 
glass in the stem, 

N = number of degrees of emergent mercury column, 

i = observed temperature, and /, = mean temperature of the emergent column. (Fig. 7.) 
Absolute Temperature. In addition to the three temperature scales already described 
physicists employ what is known as the "absolute^ scale of temperatures," based on the so-called 
"absolute zero of temperature," at which point no molecular vibration exists. This zero is con- 
ceived as 491.6® F. below the melting-point of ice, or 32** F., it having been discovered that an 

ideal perfect gas would change in volume by -r^-r-z of its volume at 32** for each 1** change in its 

491.0 



PHYSICAL UNITS AND THE MEASUREMENT OF HEAT 9 

temperature at constant pressure. Thus, if 491.6 cu. ft. of gas measured at 32" F. is cooled 20** 
F. at constant pressure the new volume will be 471.6 cu. ft. 

It is only necessary to add 491.6 — 32 or 459.6 to the actual thermometer reading to get 
the absolute temperaturei that is, T » < + 459.6, where T = absolute temperature, and t — 
actual thermometer reading on the Fahrenheit scale. For engineering work 460" is used rather 
than 459.6". For th3 Centigrade scale the relation ia T =^ t + 273.1. 

Pyrometers. For the measurement of high temperatures above 500" F. pyrometers of 
various kinds are employed. Mercurial pyrometers may be used for flue-gas temperatures up 
to 1000" F. These arc simply thermometers with an inert gas such as nitrogen or carbon 
dioxide fcMroed in above the mercury column to prevent the mercury from boiling, since at 
atmospheric pressure it will boil at 676" F. In fact, vaporization begins much below this 
temperature, so that ordinary thermometers should not be used much above 400" F. 

Expansion pyrometers made up of two dissimilar metals, such as brass and iron, are used 
for temperatures up to 1500" F. They are liable to error unless both the brass and iron elements 
are uniformly heated throughout. In the common form a brass rod is enclosed in an iron pipe 
and one end of the rod attached to a cap at the end of the pipe, while the other end is connected 
by a multiplying gear to a pointer moving aroimd a graduated dial. Lost motion in the gearing \ 
is <^ten a source of error. 

Thermo-electric pyrometers are used for temperatures up to 2900" F., and are described in 
"Steam," Babcock <fc Wilcox Co., as follows: 

"When wires of two different metals are joined at one end and heated, an electromotive 
force will be set up between the free ends of the wires. Its amount will depend upon the com- 
position of the wires and the difference in temperature between the two. If a delicate galva- 
nometer of high resistance be connected to the 'thermal couple,' as it is called, the 
deOecUon of the needle, after a careful calibration, will indicate the temperature very 
accurately. 

"In the thermo-electric pyrometer of Le Chatelier, the. wires used are platiniun and a 10 
per cent aUoy of platinum and rhodimn, enclosed in porcelain tubes to protect them from the 
oxidizing influence of the furnace gases. The couple with its protecting tube^is called an 'ele- 
ment.' The elements are made in different lengths to suit conditions. 

"It is not necessary for accuracy to expose the whole length of the element to the tempera- 
ture to be measured, as the electromotive force depends only upon the temperature of the juncture 
at the closed end of the protecting tube and that of the cold end of the element. The galva- 
nometer can be located at any convenient point, since the length of the wires leading to it simply 
alter the resistance of the circuit, for which allowance may be made. 

"The advantages of the thermo-electric pyrometer are accuracy over a wide range of iem- 
peratures, continuity of readings, and the ease with which observations can be taken. Its 
disadvantages are high first cost, and, in some cases, extreme delicacy.'' 

Tor temperatures up to 3227" F., the fusing point of platinimi, it is possible to make use 
of the melting points of various metals for approximate temperature indications. 

Above these temperatures either radiation or optical pyrometers are employed, the former 
having a range as high as 3600" F., the limit in steam-boiler practice, and the latter being capable 
of recording temperatures as high as 12,000" F. 

Measurement of Heat Quantity. (Calorimetry.) Heat may be measured^ since it is a form 
of energy, in any of the usual energy units, as the joule, foot-pound, or horsepower hour. 
However, it is the custom to use for this purpose a special unit more readily applicable to 
heat changes. This unit in the English system is known as the British thermal unit (B.t.u.), 
and is the amount of heat required to raise 1 pound of water from 63" to 64" F.; in the 
French system the unit is called the Calorie, and is the amount of heat required to raise 1 
kilogram of water from 15" to 16" C. Since 1 kg. = 2.2046 lb. and 1" C. = ^/s** F., then 1 Cal. 
» 2.2046 X V? = 3.968 B.t.u. or 1 B.t.Xi. = 0.252 Cal. 

The tendency at the present time is to define a B.t.u. as the mean or average amount of 



10 



POWER PLANTS AND REFRIGERATION 



heat per degree to raise 1 lb. of water from 32^ to 212^ F., which is almost exactly the same 

the heat required to raise 1 lb. of water 1^ at 63.5^ F. 

The calorimeter is an apparatus into which a hot body oi known temperature and weight 

can be introduced, and in cooling throu^ a known difference in temperature is made to give 

up heat measured in B.t.u. to a liquid also of known 
temperature and weight, which imdergoes a coneqxiDd- 
ing increase in temperature. 

Orvm 




A^d fr/rJ/oo 



Mech . ^iiiy. \^ \ 



\ 



L 

Temp. rise*/ f 
Heot oddest' /Rta 
Work. dm,'mn/l^i 



I 
I 
I 






L. J 



^j//fy Example. If 1 lb. of iron Is put into a calorimeter con- 

'*^ taining 10 lb. of water, and the water rises in temperature 

5® F., the iron has given up 50 B.t.u., and at the same time 

its temperature has fallen about 420^ F. 

If 50 lb. of water are raised from 70® to 90** F. it is 

customary to say that 50 X (90-70) - 1,000 B.tu. have 

been added. 

It should be noted that while a B.t.u. is based on 
the temperature interval of 63*^ to 64** F. it will be suffi- 
ciently accurate for engineering work to use the actual 
temperature interval direct in any case without correo- 
tifm into terms of 63*' to 64'' F. 

Specific Heat, It b a well-known fact that equal 
quantities of heat will raise equal weights of different 
substances a different number of degrees, depending on 
the natm*e of the substance. This property of matter 
is known as specific heatf and for any substance can be 
expressed as the number of B.t.u. required to raise or 
lower the temperature of 1 pound 1^ F. at some given 
temperature. It is also customary to make use of the 
mean or average value for a certain temperature intervaL 
Two specific heats are recognized, one known as 
the "true" specific heat, measured at the temperature 
stated, and the other as the "mean" specific heat, which 
PiQ^ 3^ is the average value between the temperatures under 

consideration. In the case of gases a further distinction 
is made between specific heat at constarU pressure and 
at constant volume. See "Specific Heat of Gases" in chapter on "Air." 

The specific heat at constant pressure of a mixture of gases is obtained by multiplying the 
specific heat of each constituent gas by the percentage by weight of that gas in the mixtiwe, 
and dividing the sum of the products by 100. The specific heat of a gas whose composition by 
weight is COt, 13 per cent; CO, 0.4 per cent; 0, 8 per cent; N, 78.6 per cent, is found as foUows: 

COi : 13. X 0.217 « 2.821 

CO : 0.4 X 0.2479 « 0.09916 

: 8. X 0.2176 - 1.74000 

N : 78.6 X 0.2438 » 19.16268 

100.0 23.82284 

and 23.8228/100 >■ 0.238 » specific heat of the gas, at constant pressure. 

The specific heats of various solids, liquids, and gases are given in Table 1. 

Relation between Units of Energy and Power. Since the various forms of energy, heat, 
mechanical energy, electrical energy, etc., are mutually convertible there must be definite numer- 
ical relations between the variovis imits used to express energy. As determined by various physi- 
cists the relation between the B.t.u. and ft.-lb. is 



• / /fP -SiOQP f/./l» perm/a 
fHP'iy^^4:!Jiatu.p&:min 



1 B.t.u. « 777.64 ft.-Ib. 



PHYSICAL UNITS AND THE MEASUREMENT OF HEAT 



11 



The number 777.64 is called the mechanical equivalent of heat and is denoted by J. For ordinary 
use the value 778 may be taken. Another convenient relation is, 1 hp.-hr. = 2,546 B.t.u. 

TABLE 1 

SPECIFIC HEATS OP VARIOUS SUBSTANCES 







80LI08 










Temperature,* 

Degrees 

FahnaoheUt 


Speeifie 
Heat 


• 


Temperature,* 

Decrees 

Fahrenheit 


Speeifle 
Bsat 


Copper 


69-460 
32-212 
5»-212 
6S-212 
68-208 
68-208 
82-212 
82 


0.0961 GlaM ^normal ther. 16">).. 


66-212 
69 

32-212 
82-212 
-106-64 




0.1988 


GewT:;..... ..:.:::. ::..:.:.. 


.0316 


Lead 


.0299 


wfpouKnt ipon . 


.1162 
.1189 


Ptatinum 


.0828 


Cast mm 


Sflver 


.0669 


Steel (aoft) 

Steel Omrd) 


.1176 
.1166 
.0986 
.0888 


Tin 


.0618 


loe 


.6040 


Zine. '.....' 

Braaa (yellow) 


Sulphur (newly fused) .... 


.2026 



UQUID8 



Water. 
Aleohol 
or: 

(mcieed) 



Temperature,* 

Degrees 

Fahrenheit 



69 
j 82 
1l76 

82 
i 60 
(122 

69-102 
to 360 



Soeeific 



1.0000 

0.6476 

.7694 

.3346 

..4066 

.4602 

' .0410 



Sulphur (melted) 

I Tfci (melted) 

Sea-water (8p.gr.l.0043) . 
Sea-water (8p.Kr. 1.0463).. 

Oil of turpentine 

Petroleum 

Sulphuric add 

Olive dl 



Temperature^* 

Degrees 

Fahrenheit 



246-297 

"64*" 
64 
32 

64-210 
68-133 



S] 



iDocifle 
Heat 



0.2860 
.687 
.980 
.908 
.411 
.498 
.3368 
.309 





Tempera- 
ture,* 
Degrees 
Fshrenheit 


Constant 
Preesure 


Specific 

Heat at 

Constant 

Volume 




Tempera- 
ture,* 
Degrees 
Fahrenheit 


Specific 

Heat at 

Constant 

Preesure 


Specific 

Heat at 

Constant 

Volume 


Air 


32-392 
6&-40& 
32-392 
64-888 


0.2376 
.2176 
.2438 

8.4090 


0.1693 


Carbon monoxide 


41-208 
62-417 
64-406 


0.2426 
.2169 
.6929 
.2277 
.2400 


0.1728 


CaOTBD .......^ 


.1663 Carbon dioxide 


.1636 


Nitiegen 


.1729 
2.4141 


Methane. 


.4606 


'1 vdrofl'tB . 


Blast^Fur. gas (approz.) 
Flue gas (approz.) 















SPI 


XHFIC HEAT OF BUILDING 


MATERIALS 




BnildiBc Materials 


SMdfic 
Heat 


Building Materials 


Specific 
Kt 


DensitiM 


Lb. 
clu Ft. 


Brickwork 


0.1960 
.2169 
.2000 
.4670 


• 
Oakwood 


0.6700 


Stone woric 


160 


Masonry 


Bfrch 


Wood 


40 




Glaas ;i977 

1 


Slate 


170 




Plaster 


90 






i 
1 







one temperature alone is given the 
for the range of temperature given. 



'true" spedfle heat is given; otherwise the value is the "mean** 



One method used for determining the value of 7 is shown diagramatically in Fig. 8. This 
apparatus consisted essentially of a paddle-wheel revolved by a cord wound around a drum 
and ooonected to a known weight which in falling through a known distance caused the wheel 
to stir up the water and thus transmit the energy of the falling weight to the paddle. The 
friction of the water against this wheel produces heat which raises the temperature of the 
water a knofwn number of degrees. 

The upit of electrical energy is the joule, and the corresponding unit of power is the wattf 



12 



POWER PLANTS AND REFRIGERATION 



or one watt is the some as one joule per second. The larger u 
1,000 watte. The following are the rektioofi between these u: 
equiBolttUa qf heal. 

1 watt-hour - 3.415 B.t.u. 

1 kw.-hour - 3,415 B.t.u. 

1 hp.-hour = 746 watta = 0.746 kw. 

The nutnerioal relations between the various units of pressure, energy, anil power is giv^n in 
the following table. ' ' 

TABLE 2 



Equlnlant Valu* in Otbir UniU 



id whli psrfeet afflelaiey 



1,980,000 fc lb. 



S.B4 lb. mCfir tn 



1.34 hone-pomr 
2,6M,Z0O ft. lb. per hour 
U.24D ft. lb. per midUM 
TST.S It. lb. per •eeoad 



vipontAJ per hour frwn u 



33,000 It. lb. per rainu 



I.T07 B.t.ii. per kcoi 
I.ITS lb. arboD oild 



Sellable and Latent Heat. 'R'henever wc add heat to a substance without change of sta(« 
wc increase its temperature, and the heat thus added is known as serui&U heal, as, for example, 
the heat added to water between 50° and 140° F. Sensible heat changes, as already stated, 
are measured by the thermometer. 

Heat may be added to a body without any change of temperature provided a ctumge <rf 
Rtate from solid to Uquid or ftoni liquid to vapor takes place, and the heat thus added is known 
as latent heat. When the change b from solid to liquid, as ice to wat«r, this heat is known as 
the latettl heat of /uaion. At atmospheric pressure ice melts at 32° F. and the latent heat is 
144 B.t.u. per pound. 

When the change b from Uquid to vapor, as water to st«am, the heat required to effect the 
change is known as the latent heat <4 evaporation. At atmospheric pressure water evaporates at 
212° F. and the latent heat is 971.7 B.t.u. per pound. 



PHYSICAL UNITS AND THE MEASUREMENT OF HEAT 



13 



TABLE 3 

APPROXIMATE MELTING POINTS OF METALS AND OTHER SUBSTANCES 



Metal or Otber 
Subctanee 



i Temperature, 
Deg . Fahrenheit 



Wroufht iron 

Pig iron (^P*y) • • • 
inn (wmte)« 



Metal or Other 
Subetanoe 



Steel (east). 
Copper. . . . 

Zinc 

Antimony. . 

lee 

TaUow 

Stearic add. 
Sulphur . . . . 



2787 
2190-2827 

2075 

2460-2660 

2500 

1981 

786 

1166 

32 

92 

158 

289 



Lead 

Bismuth 

Tin 

I^tinum 

Gold 

saver 

Aluminum 

Mercury 

Carbon dioxide. 
Sulphur dioxide , 



Temperature, 



Oeg. 



operatun 
Fahrenh< 



eit 



621 

498 

449 

3191 

1946 

1762 

1216 

-89 

-108 

-148 



In DO ca^e is this latent heat lost, as it always reappears whenever the substance passes 
through the reverse process from gas or vapor to liquid or from liquid to solid. 

The temperature of ebullition of any liquid, or the hoUing-paint, may be defined as the tem- 
perature which exists when the addition of heat to the liquid no longer causes rise of tempera- 
ture, the heat added being absorbed or utilized in converting the liquid into vapor. This tem- 
perature is dependent upon the pressure under which the Uquid is evaporated, being higher as 
the pressure is greater. See Table 5 in the Chapter cm '* Water, Steam, and Air." 

Expansion of Solids. The addition of heat, to practically all substances, causes them to 
expand or increase in length, area, and volume, providing no change of state takes place during 
heating. The amount by which one unit bf length, area, or volume of the substance changes in 
length, area, or volume per V rise in temperature is known as the coefficient of linear, superficial^ 
or cubical expansion, respectively. The coefficient of expansion is not a constant quantity and 
hence the temperature range to which the coefficient apphes should always be stated. The varia- 
tion is slight for the same material and the coefficient is usually assumed constant for any given 
substance. ^ 

TABLE 4 

LINEAL EXPANSION OF SOLIDS AT ORDINARY TEMPERATURES 
(Tabular vahiea repr ce cn t increase per foot per 100 degrees increase in temperature, Fahrenheit) 



Substance 



Brass (cast) 

Brass (wire) 

CJcppof 

Glass (English flint) 

Granite (average) 

Iron (cast) 

Iron (soft forged) 

Iron (wire) 

Lead 

Mcreuryt 

li me s t o n e 

Sleel (Bessemer rolled, hard) 

Steel (Bessemer rolled, soft) 

Steel (cast, French) 

Steel (cast annealed. English) 

* Where range of temoerature is given, coefficient is mean over range, 
t Coefficient of cubical expansion. 






_ 


Temperature 

Conditions,* 

Degrees 


Coefficient 


^_100 
Degrees 


Fahrenheit 


Fahrenheit 


82 to 212 


0.001042 


82 to 212 


.001072 


82 to 212 


.000926 


82 to 212 


.000451 


82 to 212 


.000482 


104 


.000689 


to 212 


.000634 


82 to 212 


.000800 


82 to 212 


.001605 


82 to 212 


.009984 


82 to 212 


.000189 


to 212 


.00066 


to 212 


.00063 


104 


.000734 


104 


.000608 



PropogAtion of Heat Heat may be propagated by conduction, convection, and radiation. 

Conduction is a molecular transmission of heat, the material in question transmitting the 

heat frcxn particle to particle of its own substance. This transmission will only occur between 



14 POWER PLANTS AND REFEUGERATION 

any two sections of the material which are at different temperatures, the heat alwa3r8 flowing 
from the higher to the lower temperature. 

Time is required for conduction to take place, and varies with the distance between the 
sections, with the temperature difference, and with the character of the material. Good o(hi- 
ductorsL permit a very rapid flow, while poor conductors transmit heat very slowly. In theoe 
latter substances great differences of temperature may exist, while in the ioinn& the substance 
arrives at very nearly the same temperature throughout in a very short time. 

Since conduction takes place between molecules by contact it may go on in any direction 
from the source of heat, and hence does not always travel in straight-lines like radiation. The 
amount of heat which is transmitted per unit of time by conduction is directly pn^xHticHial 
to the area of the cross-section, to the difference of the temperatures divided by the thickness, 
and to a coefficient which depends on the character of the material. 

The coefficient of condvction is the quantity of heat which flows in unit time, through a cross- 
section of "unit area, when the thickness of the plate is unity and the difference of temperature 
is one degree. In the English system the relation that determines this coefficient is 

Q « quantity of heat in B.t.u. C » coefficient of conduction per 1 in. thickness, S = area in 
sq. ft. X s thickness in inches, tt ^ k " the temperature difference between the two sections 
or surfaces, and T » time in hours. 

The conducting power of substances varies greatly, as shown by the table of abs(^ute con- 
ductivities of various materials in the Chapter on " Heat Transmission of Cold Storage Walls.'' 

Convection is the transmission of heat by the cimilation of one substance, a fluid or gas, 
over the surface of a hotter or colder body. The particles or molecules of the moving substance 
come into dose contact with the hotter body, and are actually heated by conduction during the 
period of this contact, but immediately pass on, carrying what heat they have acquired along 
with them, and fresh, cooler molecules succeed them. This circulation may be caused by purely 
natural forces, or may be product by mechanical means. The circulation of the water in a 
boiler is an example of the former, while the circulation of air over the heater coils in a fan blast 
heating system is an example of the latter condition. In case the circulating substance is hotter 
than the other body the process will be reversed and heat will be given up by the moving molecules. 

In general, it may be said that the heat transferred by convection is independent of the 

nature of the surface of the body and of the surrounding absolute temperature. It depends on 

the velocity of the moving substance, varying as some function of the velocity, on the form and 

dimensions of the body; and on the temperature difference between the moving substance and 
the body. 

The general expression for the heat given off by convtction is: 

where V ia the velocity in feet per second, t, and t^ are the temperatures of the heating media and 
outside air respectively in degrees F., A is the area in sq. ft., and / and n are constants to be 
determined for the radiator in question. 

Radiation is the transmission of heat through a medium conmionly known as the ether, 
which is assumed to occupy all intermolecular spaces. Radiation always takes place in straigHr 
lines, obeying the same laws as light, so that its intensity or amount per unit of surface varies 
inversely as the square of the distance from the source of radiation to the surface, and directly 
with the sine of the angle of inclination. Moreover, radiant heat continues to travel in the 
same straight line until intercepted or absorbed by some other body. 



PHYSICAL UNITS AND THE MEASUREMENT OF HEAT 15 

TABLE 5 

RELATIVE RADIATING OR ABSORBING POWER AT 212<* F. 



T.«^pl)ladf. , ^ ^ 


100 
100 
98 
90 
85 
72 


Steel 


17 


WbitW IfMMl 


Platinum ^ ... , 


17 


J^mptr 


PoliiA)^ hraaB 


7 


Glua , 


C<mper 


7 


IfMiift mk . . 


PoUahed gold 


8 


Shfilbu; 


" iayer 


8 









The amount of raditot heat emitted or absorbed depends largely upon the character of 
the surface of the hot or cold body, and it has been f oimd that the power of a given substance 
for absorbing radiant heat is exactly the same as for emitting radiant heat. Table 5 gives the 
relative radiating powers of various substances at 212° F. 

Radianl heat has the property of passing through dry gases without heafing them to any ap- 
preciable esctent, but air containing water vap<nr or dust will intercept and absorb radiant heat, 
henoe the earth's atmosphere is warmed by the radiant heat from the sun. 

Radiant heat like light is reflected from various materials, and it will be found that in gen- 
eral sabstajices possessing a high power of radiation have a low reflecting power. Silver has a 
relative radiating power of 3 but its reflecting power is given as 97. 

Radiant heat will also pass through certain solid substances without heating them, in the 
same way as light passes through glass. This property of substances is known as diaihermacy, 
and crystals of rock salt have this property to a very high degree. 

Radiant heat is diffused in all directions by certain materials such as white lead, powdered 
silver, and chromate of lead. Radiation in this case takes place in all directions, with little ot 
no regularity or uniformity of direction. 

In general it may be said that the heat emitted by radiation per unit of surface and per 
unit of tiine is independent of the form and extent of the heated body provided there are no 
re-entrant surfaces to intercept the heat rays. Also, the amount of heat emitted by a surface 
radiating equally in all directions depends only on the natxue of the surface, the difference in* 
temperature between the surface and surroundings, and the absolute value of the temperature. 
The general expression for heat given off by radiation, as stated by Newton, and later by Dulong 
and PetU, as well as Sttfan and BoUzman, is: 

where K is the radiation constant or coefficient, and Ti and Tt are the absolute temperatures of 
the hot body and the surroimding colder bodies respectively. Newton gave the exponent x the 
▼alue 1, but this has since been proved too small, and Stefan and BoUzman give the value x » 4» 
while for a black body they give /C = (16 X 10"*^. 



1 



CHAPTER II 

WATER, STEAM, ANJX AIR 



WATER 

Properties of Water. Pure water is a chemical compound (HsO) formed by the imion of two 
volumes of hydrogen gas with one volume of oxygen gas, or 2 parts by weight of hydrogen and 16 
parts by weight of oxygen. Water expands when heated from 39.2° F., the temperature of maxi- 
mum density, to any higher temperature, but contracts when heated from 32** to 39.2° F. At the 
atmospheric pressure of 29.92'' mercury its freezing point is 32** F. and its boiling point is 212** F. 
• The change in density is shown by the following comparison of weights per cu. ft. at various 
temperatures. 



At 32^ ¥., or freezing point 
" 39.2** F., max. density 
" 62* F., standard 
" 212*» F., or boiling point 



62.418 lb. per cu. ft. 
62.427 " " " " 
62.365 " " " " 
69.760 " " " " 



TABLE 1 

HEAT CONTENT AND SPECIFIC WEIGHT OF WATER 



Tonp.f 


HMt 

Content 


Weight, 


Temp.* 


Heat 
Content 


Weight, 


Temp., 


Heat 
Content 


Weight, 


Deg. 
Fahr. 


Above 82« 
per 1 Lb. 


Lb. per 
Cu.Ft. 


Deg. 
Fahr. 


Above 32<' 
per 1 Lb. 


Lb. per 
Cu. Ft. 


Deg. 
Fahr. 


Above 82<' 
per 1 Lb. 


Lb. per 
Cu.Pt. 


82 


0.00 


62.42 


100 


68.00 


62.02 


158 


125.88 


61.02 


35 


3.02 


62.42 


102 


69.99 


62.00 


160 


127.88 


60.98 


40 


8.05 


62.42 


104 


71.99 


61.97 


162 


129.87 


60.94 


45 


13.07 


62.42 


106 


73.98 


61.95 


164 


131.87 


60.90 


60 


18.08 


62.41 


108 


75.97 


61.92 


166 


133.86 


60.85 


52 


20.08 


62.40 


110 


77.97 


61.89 


168 


135.88 


60.81 


64 


22.08 


62.40 


112 


79.96 


61.86 


170 


137.88 


60.77 


66 


24.08 


62.89 


114 


81.96 


61.83 


172 


139.88 


60.73 


68 


26.08 


62.88 


116 


83.95 


61.80 


174 


141.88 


60.68 


60 


28.08 


62.87 


118 


85.94 


61.77 


176 


143.89 


60.64 


62 


. 30.08 


62.36 


120 


87.94 


61.74 


178 


145.89 


60.59 


64 


S 32.08 


62.35 


122 


89.93 


61.70 


180 


147.89 


60.66 


66 


34.08 


62.34 


124 


91.93 


61.67 


182 


149.89 


60.60 


68 


36.08 


62.33 


126 


93.92 


61.63 


184 


151.90 


60.46 


70 


38.07 


62.31 


128 


95.92 


61.60 


186 


153.90 


60.41 


72 


40.07 


62.30 


130 


97.91 


61.56 


188 


156.91 


60.37 


74 


42.07 


62.28 


132 


99.91 


61.52 


190 


157.91 


60.32 


76 


44.06 


62.27 


134 


101.90 


61.49 


192 


159.92 


60.27 


78 


46.06 


62.25 


136 


103.90 


61.45 


194 


161.92 


60.22 


- 80 


48.05 


62.23 


138 


105.90 


61.41 


196 


163.93 


60.17 


82 


60.05 


62.21 


140 


107.89 


61.37 


198 


165.94 


60.12 


84 


62.04 


62.19 


142 


109.89 


61.34 


200 


167.96 


60.07 


86 


64.04 


62.17 


144 


111.89 


61.30 


202 


169.95 


60.02 


88 


66.08 


62.15 


146 


113.89 


61.26 


204 


171.96 


59.97 


90 


68.03 


62.13 


148 


115.88 


61.22 


206 


173.97 


59.92 


92 


60.02 


62.11 


150 


117.88 


61.18 


208 


176.98 


59.87 


94 


62.02 


62.09 


152 


119.88 


61.14 1 


210 


177.99 


59.82 


96 


64.01 


62.07 


154 


121.88 


61.10 1 


212 


180.00 


59.76 


98 


66.01 


62.05 


156 

1 


123.88 


61.06 









At 62® a U. S. gallon of 231 cu. in. weighs approximately S}4 lb., and a cu. ft. is equal to 
7.48 gals. Pressures are often stated in feet or inches of water column, and at 62^ F. the equiva- 

16 



WATER, STEAM, AND AIR 



17 



lent in pounds per sq. ft. is (let h = head in feet), = 62.355 A, or in pounds per sq. in. 
62.355 



144 



k = 0.433 h. Also, if hi = head in inches of water at 62° F., then the pressure in ounces 



per sq. in. = '^ — X 16 = 0.578 /ii, or hi = 1.73 X pressure in ounces per sq. in. A column 

12 -144 

of water 2.309 ft. or 27.71 in. high exerts a pressure of 1 lb. per sq. in. at 62** F. 

For density of water at other temperatures than those already stated see Table 1. 

The specific volume of water, or the volume of one poimd, depends on the temperature at 
which the volume is measured, and is practically independent of the pressure, since water is but 
very slightly compressible. The specific volume is the reciprocal of the specific density, 
values for the latter being given in Table 1, hence it is only necessary to find the value of 

— to get the volume of 1 pound, as tt-tz =» 0.016 cu. ft. at 32* F. 

wt. per cu. ft. 62.42 

The bailing point of pure water varies with the pressing or altitude above sea level, the tem- 
perature at which ebullition will occur decreasing wiih the altitude or lower pressure. This 
relation is shown by reference to the steam tables, which also indicate that the boiling point 
increases for pressures higher than that of the atmosphere at sea level. Table 2 gives the 
boiling points at various altitudes. 

TABLE 2 
BOIUNG POINT OF WATER AT VARIOUS ALTITUDES 



Bnainc 
Foint, 

Decreee 
Fahr. 


Altitude 
Above 


Atmospheric 
Prenure, 


Sea 
Level, 


Pounds 


Feet 


Sq. In. 


184 


15.221 


8.20 


185 


14,649 . 


8.88 


186 


14.075 


8.67 


187 


13.498 


8.76 


188 


12.934 


8.95 


189 


12.367 


9.14 


190 


11.799 


9.34 


191 


11.243 


9.54 


192 


10.685 


9.74 


198 


10.127 


9.95 


194 


9.679 


10.17 


196 


9.031 


10.39 


196 


8.481 


10.61 


197 


7.982 


10.83 


198 


7.381 

1 


11.06 

> 



Barometer 


Bofling 
Point. 

Decree 
Fahr. 


Altitude 


Atmospheric 


Barometer 


Reduced 


Above 


Pressure, 


Reduced 


to 32 


Sea 


Pounds 


to 32 


Degrees. 


Level. 


per 


Degrees* 


Inches 


Feet 


Sq. In. 


Inches 


1 
16.70 


199 


6343 


11.29 


22.99 


17.06 


200 


6.304 


11.62 


23.47 


17.45 


201 


5.764 


11.76 


23.96 


17.88 


202 


5.225 


12.01 


24.46 


18.22 


203 


4.697 


12.26 


24.96 


18.61 


204 


4.169 


12.61 


26.48 


19.02 


205 


3.642 


12.77 


26.00 


19.43 


206 


3.115 


13.03 


26.63 


19.85 


207 


2.589 


13.30 


27.08 


20.27 ! 


208 


2.063 


13.67 


27.63 


20.71 


209 


1.539 


13.85 


28.19 


21.15 


210 


1.025 


14.13 


28.76 


21.60 


211 


512 


14.41 


29.33 


22.05 


212 


Sea 


14.70 


29.92 


22.52 

1 




Levd 







The specific heat of water, or the number of B.t.u. requij^ to raise the temperature of 
1 pound of water 1® F. varies with the temperature as shown in the following table. 



TABLE 3 



Tonperature. F.** 

30OF. 

66 
100 
160 
210 



Specific Heat 
1.0098 
1.0000 
0.9967 
1.0002 
1.0060 



In consequence of this variation, the amount of heat required to raise 1 lb. of water at 32^ F. 
throu^ a known temperature interval, known as the heat of the liquid^ will depend on the average 
value of the specific heat for that range, and this variation is shown in Table 1 — ^where the 
'lieat units" required to raise 1 lb. of water from 32^ F. to the temperature in the table is 
given as the heat content. 



Jg POWER PLANTS AND REFRIGERATION 

The specific heat of water is veiy commcmly Bsaumed to be unity, and ia so used in many 
engineering caiculationa. The steam tables, however, are based on the exact value [or the 
temperature range in question. 

The specific heat oF ice at 32° F. is 0.463 B.t.u. per 1 pound. 

Flow of Water in Pipes. The flow of water in pipes depends on a difference in head or 
pressure between the two points between which flow takes place. Tliia difference in head is used 
up in overcoming the resistance (friction of the pipe) leered to the flow, and in creOitiiig the 
velocity of discharge at the second point. 

The flow of a liquid in a pipe is under the influence of three heads or equivalent prcAuree. 

The vdocUy head or prttmre is defined as that bead or pressure of the liquid which is required 
to create the velocity of flow, that is, the head or pressure necessary to accelerate the mass from 
a state ol rest to the velocity attained at the point in the line under consideration. 

The reiufonce head or pneture, also termed the fndiim head, is that head or prenure required 
to overcome the frictional resistance offered to the flow. 

The Uital head or pressure, also termed the dynamic head or pressure, ia the sum of the velocity 
head plus the friction head. 

The potential or measured head is the vertical distance, measured in feet, from some datum 
line to the center of the pipe at the point in the line under conaideratioD. ' 

Hie Piezometer in its simpleat form consists of a tube inserted in a pipe at right angles to 
the flow (Fig. 1). The radial pressure within the pipe is measured by the height of the column 
of the liquid within the tube. For high pressures an ordinary gage of the Bourdon type ia 
substituted for the tube. 

The reading obtained by the use of a pieiometer placed in a pipe of uniform cross sectica 
throughout its entire length with free discbarge to the atmoephert, ia the head lost by friction be- 
yond the point of attachment of the piezometer. 

Hw Pitot Tube in Its simplest form is a bent tube placed in the pipe so that the immwsed 
end of the tube faces the stream (Fig. 2}. The hei^t of the column of liquid in the tube is greater 



than the reading obtained by the plesometer by an amount equal to the head required to produce 
the velocity of flow. The height of the column is the UAal head at the point of measurement. 

The difference between the reading obtained by the Pitot tube and the piesometer is the 
velocity head at the point considered. If the pipe is of uuifimn diameter this difference is of 
course a constant throughout the length of the pipe as the velocity is constant. 

The difference between the total and resistance heads is read direct on the manometer bjr 
connecting the opposite ends of the U tube to the pieiometer and Pitot tube as shown by Fig. 2. 

If it were not for friction "the total head at any point or section would be equal to tlw toUl 



WATER, STEAM, AND AIR 19 

head at any Bubeequent point or section," total head being the sum ol the static or fricticoi head 

plus thp velocity head. See Fig. 4. H = k, + h,, where 

H — total bead measured in feet ol fluid flowing. | jj 

A, = static or friction head measured in feet of fluid tf l~ '^^^^ 

flowing. j i — __^^^^*''*^ 

h, " velocity head meOBUred in (eet of fluid Bowing. L ^^''■'"■Vs.^^^V , 

This rel .tion between the total head at any two sections z, Bcmajlfja ^^Sj; 

d a pipe line, aaauminR frictionlexs flow, is known as Ber- I ThoorWTt ' /^^ 

fumUt'« Uuorrm and is demonstrated as foilowH. 3ee Fig. 3. ?' ^ 

Assume (Ij a perfect fluid, (2) steady flow, (3) no frio- - — — *— 

tioD. Aaume a wei^t W paaaes section A in unit time. pio^ 3, 

Becauae of (2) a wei^t W also passes B in the same time. 



Kinetic energy of If at .A = ^MV** f ' 



where M — — 1 



Let Pi » the radisJ or static pressure at section A measured in pounds per sq. ft. and 
Pi the static pressure at B ibeasured in pounds per sq. ft. 
p, 
PotenUal energy of IF at A = WZi + -rWia which D is the density of the fluid flowing hence 

-^ is the potential botd equivalent to the static pressure Pi, and Zi = potential, bead or measuied 



head ftt the section. The total energy of If at A - (cf—^ +Z, + —M. likewise, the total 

ranKv of IT at fi - Wi — + Zt H — -]. Smce there is no external frictional resistance the 
\2j DJ 

total energy at A equals that at £ or 

2g D 2g D 

Tliia is Bernoulli's theorem, and each member of the equation is the "total head" at the 
eorreqianding section. It may be stated thus: In a steady flow mthmit frieUon ihA total head 



. KiBMk MWSJF of tf - H u v,< - - 



20 



POWER PLANTS AND REFRIGERATION 



at any section equals the total head at any subsequent section. Note that the ''total head" is 
the sum of the "velocity" head, the "potential" head, and the "pressure" head. 

Case L (Flow withotU friction): Apply Bernoulli's theorem to the case of water issuing 
from the base of a stand pipe. See Fig. 1. 

The pressure at A is atmospheric (Pa) and within the jet at B it is also atmospheric. The 
velocity at A is zero 

p yi p 

+ H + -^ = . ^ . /- 



:.H 



2g 



2g ^ ^ D 
or F = '^2gH 



Case IL (Flow with friction) : Since friction tends to oppose motion the total head at any 
section is greater than the total head at any subsequent section. The "lost" or "friction" head 



r 



^^n: 



RasQnfoIr 







f-f. 



^ 



ffn 



P.^\^Su, 



__i 



wp. 



Pio. 6. 



between any two sections is therefore the diflfercnce between the total heads at these sections. 

(Fig. 4.) 

P 
The total head atA=0 + //-f-yj- 



The total head at C 



1/2 p 

2^ ^ ^ D 



The friction or lost head = (o+^+§)-(t- + 0+~) =H-^ 



WATER, STEAM, AND AIR 21 

By appl3ring the equation between A and B, or B and C it can readily be eeen that the ''pressure" 
or "static" head at B equals the friction or lost head caused by the pipe line. 

Application of Bernoulli* 8 theorem to the case of a pump to show what the suction and dis- 
charge gages on the pump register, and to show how the "total" head on a pump may be found. 
(Fig. 5.) Call Ht the ''total" head on pump, ^i the friction head lost in the suction pipe, ht 
the friction head lost in the discharge pipe, Vg the suction velocity, and Vd the discharge velocity. 

Applying the equation between the surface of the sump and the suction gage: 



(o + o^5).^..(|l'-f^,-ff) 



The pressure registered on the suction gage is (Pa ~ P^*) where P« is the absolute pressure 
at this section. From equation W f: (Pa — P*) *= ( i/, + -r-^ + hij. 

The head registered on the suction gage equals the suction lift, plus the suction velocity 
head, plus the suction friction head. 

Applying the equation between the discharge gage and the end of the line: 

.'. -^ {Pd - P.) « /«. + i/L 

The discharge gage registers the friction head in the discharge line, plus the measured lift 
of 'the discharge line. 

Apply the theorem between the surface of the sump and the discharge pressure gage. 
(Note that the head Ht is added to the water during its passage through the pump.) 



Yli 

2g 



(2) 



.'. I — — —1 — Ht — hi — Ho — H, — 

.'. ffr = Ai + *. + H.+ Hd+ Hl + T^ 

The total head on the pump is equal to the entire friction head plus the measured head plus the 
final discharge velocity head. 

When the head is produced on the pump by closing the discharge valve, the measured head 
does not exist in reality but only virtually. The total head must be foimd from the two gage 
readings, the velocities in the suction and discharge lines, and the distance between gages. 

From equation (2) 



'"-e-5)-(«.-'-i^)-^-S'-». 



(Pi ^\ + (^ ^^ -I- i!i n.H 

\D - d)^\D- d) + 27 2^ + "" 



22 POWER PLANTS AND REFRIGERATION 

The total head equals the discharge pressure head plus the suction pressure head phis the 
final velocity head minus the suction velocity head plus the distance between gages. 

If the size of the discharge pipe equals that of the suction pipe the total head is found more 
easily. 

Vg will equal F, 

Substituting —-^ for*— — in (2) 
2q 2g 






The total head on the pump equals the discharge pressure head plus the suction pressure head 
plus the distance between gages. 

Friction Head due to Flow of Water in Pipes. The flow of waUr in a pipe of uniform diameter 
will take place with a constant velocity if the total head producing flow is maintained constant. 
This total head can be determined for any given velocity of flow -if the friction head is known. 

The loss of head due to friction when a fluid such as water, steam, air, or gas flows through 
a straight tube or pipe is generally represented by the formula, 

where/ — the coefficient of friction, L » length of tube in feet; R « perimeter of tube in feet, 
A = area in sq. ft., v « velocity of flow in feet per sec., and h * friction head in feet of the fluid 
flowing. 

ir D L t^ 4 L »■ . 

If the tube is round and D — diameter in feet, then h ^ f ^ -" T~ ^ f "T" r~ in which 

'' IT U^ 2g D 2g 

4~~ 
/ « .00644 according to Weubachy for clean iron pipe. 

2 L ff^ L t/^ 

This formula may be reduced to h — f —r- — or A = /i tt ^T" ^^ which /i — 0.02, an aver- 

D g D 2g 

age for water. 

It is understood that the pipe is smooth, clean and free from the burrs as ordinarily left by 
a wheel pipe cutter. 

For very low velocities, as found in gravity hot water heating systems, the above formula 
does not hold good. 

WiUiam Cox in the ^'American Machinist,'' Dec. 28, 1913, gives the following modification 
of the above fcurmula, which is simpler and gives almost identical /esults. 

L (4t;« -f 5i; -2) 
d 1200 

(4 1>» + 5 tJ - 2) 
Values of the expression — can be tabulated for varying velocities so that h may 

be readily solved for when v, L, and d are known. See Table 4, for these tabulated values. In 
Cox's formula d » diameter in inches. 



WATER, STEAM, AND AIR 



23 









TABLE 4 










- 


VALUES OF^^^ ^ 






• 


V 


0.0 


«0.2 


0.4 


0.6 


0.8 




1 


0.0068S 


0.00818 


0.01070 


0.01858 


0.01668 




2 


.02000 


.02868 


.02768 


.08170 


.08613 




8 


.04083 


.04680 


.06108 


.06658 


.06280 




4 


.06888 


.07468 


.08120 


.08808 


.09513 




6 


0.10250 


0.11018 


0.11808 


0.12620 


0.18468 




e 


.14888 


.15280 


.16163 


.17103 


.18080 




7 


.19068 


.20118 


.21170 


.22268 


.22368 




8 


.24600 


.25668 


.26863 


.28070 


.29818 




9 


0.80583 


0.81880 


0.83203 


0.84653 


0.35980 




10 ' 


.87888 


.88768 


.40220 


.41703 


.43213 




11 


.44760 


.46818 


.47903 - 


.49620 


.51168 




12 


.62888 


.54680 


.66268 


.68008 


.59780 




18 


0.61688 


0.68418 


0.65270 


0.67168 


0.69068 




14 


.71000 


.72968 


.74968 


.76970 


.79018 
.89630 




15 


.81088 


.83180 


.85303 


.87468 




16 


.91888 


.94068 


.96320 


.98608 


1.00913 




17 


1.06260 


1.05618 


1.08003 


1.10420 


1.12868 




18 


1.16888 


1.17880 


1.20363 


1.22908 


1.25480 




19 


1.28068 


1.80718 


1.83370 


1.86063 


1.38768 




20 


1.41600 


1.44268 


1.47063 


1.49870 


1.52713 




21 


1.55688 


1.58480 

1 


1.61403 


a. 64368 


1.67380 





The use of the formula and table may be illustrated as follows: 

Example. Given a pipe 5" in diameter and 1,000 ft. long, with 40 ft. head, what will be the di»- 
eharce? 

ifi 60 
If the velocity v is known in feet per second, the discharge will lie ir -7 X —77 X » ■* 0.32725 dh 

4 144 



cu. ft. per min. ■• Q. Now — - 

L 



hd 49 X 5 4r» + 5i>-2 



0.245 and by reference to the table it will 



1000 1200 

be seen that the actual velocity v -> 8 f t. per sec. 

The discharge in cu. ft. per min., if is velocity in feet per second and d the diameter in inches is 
0.32725 dh, hence Q - 0.32725 X 26 X 8 - 65.45 cu. ft. per min. _ 

The velocity due to the head, if there were no friction, is 8.025 V A ■> 56.175 ft. per sec. and the 
discharKe at that velocity would be 0.32725 X 25 X 56.175 -> 460 cu. ft. per min. 

Example. Suppose it is required to deliver this amount, 460 cu. ft., at a velocity of 2 ft. per sec.; 
what diameter of pipe of the same length and under the same head will be required and what wUl be the 
lose of head by friction? 



diameter 



i Q 460 , . , .. 

'\ ^^^^, - \ ^^^ - V 703 - 26.5 inches diameter. 

y V X 0.32725 \ 2 X 0.32725 



.32725 \ 2 X 0.32725 

Since the diameter, velocity and discharge are now known the frictiou head is found from 

L (4i>« + 5»— 2) 
fc -• — X — T using the table; thus. 



1000 ^^^ 20 



- 0.76 ft. 



Friction Pressure Loss Chart for Flow of Water.* The chart (Fig. 6) from the "American 
Machinist/' is based on the preceding formula. It gives the velocily of flow in pipes of various 
nominal diameters, and also ihe/riction or pressure loss in pounds per sq. in. per 100 ft. of pipe, at 
vaiying rates of flow, stated both in gallons per min., and in cu. ft. per min. 

Hie corresponding velocity of flow in lineal feet per sec. is read from the same chart by 
referring to the velocity lines, which in the example given on the sheet would be 5.9 ft. per sec. 
• For additioiial data in this eonneetkm, see the Chapter on *'Pumps." 



POWER PLANTS AND REFRIGERATION 



•- Hm-fauo aooog 

FRICTION PRESSURE LOSS LBS. PER Sa IN. PER 100 FT. 
Flo. 6, Fi/jw OF Wateh in Pipeh. 

60 gats, per min. to be transmitted 300 (t. Utrough a 2" Btiiodard Ht«el pipe. Required 
the friction loss. From 60 gals, on the left trace horiiontally to the interaection with the diagunal 
2" pipe, and read 3.25 lb. per sq. ia. at the bottom of the chart. The losa is then 3 X 3.35 = a.T.'i 

Approximait AUoaance for EUt and Olabe Vatta. 

Add to the measured leneth of liae 40 diams. for each W)° ctl, nnd 00 dinm<. f'>r carh alobe valve. 



WATER, STEAM, AND AIR 25 

Loss of Head by Entrance, Elbows and Valves. The loss of head occasioned by entrance 

F* 
to a pipe and various obstructions may be stated as a function of the velocity head as /i — — 

in which is a coefficient experimentally determined. 

YoLxiM of 0. This may be taken to equal 0.50 for a pipe at right angles to the reservoir where 
the pipe is flush with the inside surface with the burr removed so that the edge is sharp. Ap- 
proximately the same conditi(Hi exists when a smaller branch pipe is taken ofF a main. 

When Uie pipe projects inside the reservoir for a length equal to several diameters the value 
of 4> may be taken as 0.93. If the entrance is bell-mouthed and smooth the value of may be 
practically equal to 0. 

The value of 4> for elbows as stated by Weusbach based on experiments conducted with IM* 
inch pipe are as follows: 



o 



Angle of elbow = 223^** 45** 90 

Valueof0 =0.038 0.181 0.984 

For smaller pipe the value of increases. For example, Weishach gives = 1 . 53 for a 90 
degree ^inch elbow. For larger pipe the value of becomes less. 

llie value of for a globe valve, wide open, is ordinarily assumed as 1 . 5 times the value for a 
90 degree elbow. The loss through a gate valve, wide open, is ordinarily neglected. 

Elngineers, in practice, frequently assume an equivalent length of straight pipe to allow for the 
loss occasioned by elbows and globe valves. The assumption that is frequently made is to 
add to the measured length of line a length equal to 40 diameters of the pipe for each 90 degree 
^bow and 60 diameters for each globe valve. 

For further data on the loss through fittings, etc., and the allowable velocity of water through 
pipes see the Chapter on " Pumps." 

Example. A 2-inch pipe 300 ft. long with five-90** elbows and • two globe valves is to carry 60 
gaUona per min. Required the pressure loss in the line. 

From the chart Fig. 6 we find that the velocity will be approximately 6 ft. per sec. and that the 
friction loss in the straight run of pipe will be 3 X 3.25 -> 9.75 lb. per sq. in. This is equivalent to a 
head of 9.75 X 2.3 or 22.4 feet. 

6' 
The loss through 5 elbows is 5 X 0.984 X » 2.75 ft. 

2g 

0« 
The loss through 2 globe valves is 2 X IH X 0.984 X - = 1.65 ft. 

2a 

6' 
The loss of head at entrance is 0.50 X — = 0.28 ft. 

2a 

The total estimated loss of head is therefore 22.4 + 2.75 + 1.65 + 0.28 « 27.08 ft. 

Measurement of the Flow of Water. The weight of the liquid deUvered in a unit of time 
may be determined either directly or indirectly. To determine the weight delivered directly, it is 
necessary to use weighing tanks and scales or to measure the volume delivered in a tank of known 
dimensions. In the latter case the density of the liquid, by which the volum^ is multiplied to 
obtain the weight, must be known. Owing to the large size of tanks necessary when the 
quantity discharged is considerable direct measurement is frequently impractical. The indirect 
methods of determining the weight of liquid delivered depend upon the use of weirs, orifices, 
meters, Pitot tube and the Venturi tube. 

The V-Notch Weir. The apparatus consists of a tank divided into two chambers by a divid- 
ing sheet a^ shown by Fig. 7. A 90** V-notch weir is inserted in the top of the dividing sheet. 



26 POWER PLANTS AND REFRIGERATION 

Behind the weir ia the BO-caUed surge chamber or tumbling bay. The tumbUng bay is provided 
with a hook gage with scale and vernier as shown. The reading on the scale is noted witen the 
point of the hook is on the level with the bottom of Uie V-notch. A reading is made, after the 



flow starts, by raising the gage until the point of the hook begins to pierce the surface of the 
water. The difference between the two readings gives the head producing the flow over the weir. 
The formula for the 90° V-notch weir as stated by Profaaor Jamti Thompion is: 



mft. 

Where possible to adopt in practice, the V-notch weir will give consistent results and b quits 
eoctensively used in connection with the open type of feed-water beater. A recording device is 
readily attached to this apparatus through the medium of a float placed in a well which is in 
communication with the tumbling bay. 

Hie Ventttri Tube. For the racasurement of flow in pipes under pressure the Venturi tube 
(Fig. 8) is a reliable form of meter and is extensivel}' used in practice where accurate and con- 
sistent results are desired. 

The bead or preseUTe difference H between A in the " up-stream " portion of the contnicled 
tube and B at the throat is made use (rf in determining the vdocity at the throat. 

V - -^JU^ V^H CD 

■\Ia-.-au 



y = velocity at the throat, ft. per hw. 

A„ = area of " up-stream " section of tube sq. ft. 

Ai, '^ area of " throat " section, sq. ft. 

H •■ difference in head measured in ft. oi water rohinn by the 



WATER, STEAM AND AIR 



27 







■ 














^ 




c?» 














\ 












• 


• 

• 




\ 








\ 






1 






CO 


^t" 




\ 








\ 

1 












* 

«5> 






\ 










\ 












• 

Co;* 








\ 








\ 












4: 








\ 








\ 


V 




» 






1 


CO 






\ 


L 








\ 




1 






is 










\ 








\ 


\ 








CNl 


< 








\ 










\ 








<^ 










> 


\ 














* 




CV4 

( 










\ 


^ 

< 


d 
:> 


^ 

< 

■h 


J; 


p99f^ 


SI 

■ 


«: 


i 












> 


\ 




















Discharge Curves 
for 

90'' Y'Notch Weir 

Calculated from 

Thomson's Formula 




\ 


\ 






























>h 












\ 


\ 




























N 


\ 












































I 












'U-'^im U0PB9H 






28 



POWER PLANTS AND REPTIICIERATION 



The quantity of water discharged is, 
^ Q = At V cu. ft. per sec. C2) 

The velocity as determined by the above formula gives results within 3 per cent of tiie 
correct value. For extreme accuracy the meter should be calibrated by actually weighing tlie 
water for different rates of flow. 

Measurement of Flow by Means of the Pitot Tube. As previously shown the Pitot, tube 



inf0t 



Pipei to UoMcmetef 




IgggoBBBazaim 



'fmmigj^ 



lWlM-4-4 



iJ 



QaVct 




?/jnonietef 



Pro. 8. Venturi Tube with Indicattnq Manomrter. 



indicates the total pressure at the point of measurement. If a Pitot tube be placed at the discharge 
end of a pipe the reading obtained is the velocity head at the center of the pipe. 

It is a well-known fact that the velocity is greatest at the center of the pipe and least at the 
walls. The ratio between these velocities being approximately two to one, for accurate work a 
traverse of the pipe should be made, ds described in the Chapter on ** Hot Blast Heating," * and 
the relation between the velocity at the center and the mean velocity established. 

The traverse velocity curve approximates quite closely an ellipse. The mean average 
velocity is very nearly equal to 0.84 X the velocity as determined from the reading taken at the 
center of the pipe. 

Let ^ V s the velocity head measured at the center of the pipe in fe^"^ of water. 
V = velocity at center of pipe in ft. per sec. 
Vm = mean average velocity, ft. per sec 



Then V^ = 0.84 V = 0.84 \'2gh^ 



STEAM 



Properties of Steam. Steam b water vapor, which exists in the vaporous condition due to 
the fact that sufficient heat has been added to the water, from which the steam has been formed, 
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. A change in pressure will always be accompanied by a change in the temperature 
at which ebullition or boiling will occur, and there will be a corresponding change in the latent heat. 

The properties of steam together with other characteristics, are tabulated in the steam tables. 
See Table 5. 

Steam in contact with the water from which it has been generated is known as aatwraled 
steam, and may be known as dry saturated steam, or as u^t saturated steam. The latter contains 
more or less actual water in the form of mist or "priming** as it is called. 

If dry, sattu*ated steam be heated, and the pressure maintained the same as when it was 
vaporized, its temperature will increase and it will become superheatedf that is, its temperature 
will be higher than that of saturated steam at the same pressure. 

* Volume i. 



WATER, STEAM, AND AIR 



20 



A conception of the relation between the properties or characteristics of steam, and the 
manner in which the changes in state, temperature and pressure are brought about is shown 
in Fig. 12 and described in the following paragraphs. 

Generation of Steam. Consider a frictionless cylinder, Fig. 9, containing 1 lb. of water at 
32*^ F. Also consider the pressure of the atmosphere to be 14.7 lb. per sq. in. and to be replaced 
by that of the piston B, When heat is applied to the cylinder the temperature of the water 

I / 
/ 




Diagram Shomng \, >^ / 

fhe U(?u/d and iSaturafecf / 

oj^^ 



Flo. 9. Fio. 10. Fio. 11. 

I 







4 



fli 



V 




r -——zr 



.----iJi V 



\Meofadded 3f:u 




fc::r.-.:|::^-.-^-.-/fc.-.-r.-. 



1=^^ 



^"T" 



U.„._.' ff 



^c/i,'l) 



1 



FiQ. 12. 



/Vor£:- 

External i^orkofeMponftkyt - Pfi^-cfJ • Pu 

or in Bf-.u • APCy-d) ' /IPu & A - jf^ 

5/nce r ' hM hsaf^ of ^k^ap fhen r - A fii ' jnf&rno/ 

heaf' ofeyap. • (o /lecesjary h overcome molecuhr affmd^/on 



M 



Fro. 13. 



until the b<Mling point, 212^ F., is reached. The heat necessary to raise the temperature 
from 32^ F. to the boiling point is known as the ''heat of the liquid" or "sensible heat," 
and is denoted by the ssrmbol q. This condition is denoted on Fig. 12 by the point C. The 
average specific heat of water between 32^^ F. and 212^ F. is 1, hence the number of British thermal 
units (B.t.u.) necessary to raise the temperature of the water this amount is 212 — 32 or 180 B.t.u. 

When more heat is added the water begins to evaporate and expand at constant tempera- 
ture until, as in Fig. 10, the water is entirely changed into steam. This condition is also shown 
on Fig. 12, by the point D. The heat thus added is known as the ''latent heat of evaporation" 
and is denoted by the symbol r. This heat r is subdivided into two parts. See Fig. 13. First 
the attraction between the molecules must be broken down. This is known as the internal latent 
heat and is denoted by the s3rmbol p; next the external resistance must be overcome, the weight 
P being raised against gravity as in Fig. 13. The heat thus added is known as external latent 
beat and is designated by the s3rmbols APu^ where u is the change in volume, in cu. ft. of one 
pound of water, A is 1/778 and P is the pressure of the atmosphere in pounds per sq. ft. (baro- 
metric pressure). It is evident then that the latent heat r *^ p •\- APu, or p = r — APu. The 
term it Pu is the heat equivalent of the work performed for the change in volume from water to 
steam. 



30 POWER PLANTS AND REFRIGERATION 

The heat added from the starting point (32^ F.) is known as total heat (H) org + r ^ H. 
If more heat is added, the pressure remaining constant, the temperature of the steam rises and the 
steam becomes what is known as superheated steam. The heat added is equal to the mean specific 
heat (Cp) of the steam times the diange in temperature ((« — 212). Specific heat of steam is 
the B.t.u. or heat required to raise the temperature of 1 pound of the steam 1^ F. Since the 
specific heat of steam is less than that of water, the slope of this line becomes greater than that of 
the water line. The point is now located at tg on Fig. 12, and the steam has increased in volume 
in the cylinder of Fig. 10 imtil the piston occupies the dotted position B'. 

If instead of the above condition of pressure, additional pressure be added as shown by 
the weight W in Fig. 11, the temperature of the boiling point will be raised from the temperature 
of 212® F. to Bome other point aatiin Fig. 12. As may be seen by this figure, the sensible heat 
q has been increased to qi. When more heat is added the water is evapcnrated at the temperature 
ti and if heat again be added the satiurated steam will become superheated. 

Quality. The pr<^x)rtion of the dry steam per pound of steam delivered by the boiler Is 
known as the quality of the steam and is represented by the symbol x, and the heat (Hx) coo- 
tained in the steam above 32® F. is g + ar and the state point is located at ^ in Fig. 12. 

The volume of a poimd of steam is known as the specific volume (v), and, as may be seen 
by comparing Figs. 10 and 11, decreases as the pressure increases. The reciprocal of this or 

1 
weight of steam per cu. ft. is known as the density and is denoted by d or — . 

V 

The relation between pressure and specific volume for dry saturated steam is given by the 
experimental equation (Goodenough) as pv ^'^^^ « 484.2 in which 

p B pressure in pounds per sq. in. 

r = specific volume. 

Another quantity known as entropy is made use of in calculations relating to steam ftnginflw 
and turbines, and is defined as the ratio obtained by dividing the quantity of heat added to a 
substance by the absolute temperature at which it is added. The entropy of the Uquid is rep- 

r 
resented by s' or n, the entropy of vaporization by -=- and the entropy of the vapor s" or s. The use 

of entropy is explained imder the 'Hankine Cycle,'' in the chapter on ''Steam Engines." 

The total heat {H) of a dry saturated vapor for any pressure and temperature is the sum 

of the heats required to raise the temperature of one pound of the liquid from the freezing point 

to the given temperature and corresponding pressure and entirely vaporize it at this pressure. 

For this case a; = 1, consequently H ^ {p-\- APu) -^ q ^ r -\- q) H' ^ 1151.7 + 0.3746 (t - 212) 

- 0.00055 (JL - 212)«, as stated Jby Marks and Dams, 

The total heat {Ux) of wet vapor at any pressure and temperature is the sum of the 

heats required to raise the temoerature of one pound of the liquid from the freezing point to 

the given temperature and corresponding pressure and to vaporize the part z at this pressure. 

For this case, 

H, - ar + g. 

It is manifestly incorrect to say this is the heat in the vapor as the ^i Pu is not heat in the 
vwpaty but the external work performed by the vapor while evaporating. 

Heat Content of Saturated Steam. This by definition is t" « g + P + APv" in which o" 
is the specific volume of the steam. 

The total heat of saturated steam by definition is H " g + p + AP (v" — v'), in which v' 
is the specific volume of the liquid. 

As t^' is small compared with v" the term Apv* may be neglected, except at very high tem- 
peratures and pressures, and i" and H may be considered equal. 

In recent steam tables the values of i" instead of H are usually tabulated. 

Superheated Steam or Vapor. Superheated steam is defined as water vapor which has been 



WATER, BTEAM, AND AIR 



31 



bested, out of oc»tact with its liquid, until iu temperature is higher than that of mturateJ 
vapor at the uun« pressure. Moreover, if the temperature or degree erf superheat ia far 
remored from the t«mpenitur« (rf saturation the superheated vapor will follow the laws of 
perfect gasM quite ckwely, (_PV = MRT), except at high prcwuree or low temperatures. Sec 
"Air and Other Gaaea." 

The relatimt between pressure, volume, and temperature, experimentally determined 

for supertieated steam is V + 0.256 = 0.5062 — which Linde givee as a rough approximation, 



V — specific volume. 

T •- absolut« temperature. 

p •• pounds ptx sq. in. 

Tlie «iieci;Ec Aeol o/ superfceoied tJ«am ia not constant as shown by the experiments (^ K-noblaucA 
and Jakob, and others. Curves of mean specific heat are shown in Fig. 14. For any degree <£ 
sl^)erfaeat the mean q)ecific heat between the saturaticm state and the given state is given by 
the ordinate ctHresponding to the pven degree of superheat and the given pressure. For example, 
■e a! 150 lb. per sq. in. abeolule the mean specific heat for 240° superheat is 0.629. 



Hw he&t content of supahealed steam or vapor may be expressed by the equation H,— q + r 
+ f^f Ci — ') " B +Cp ((, — t) where (, — temperature (rf superheated vapiH' and ( ~ tempera- 
ture <rf saturated vapcv at the corresponding pressure, q — heat of the liquid at 1, and r — heat 
of vaporisation at tonperature t. C^ •= mean specific heat of superheat«d vapor, H — total heat 
<d one pound of dry saturated steam, and H, — total heat of one pound of superheated steam. 

Throttlijlg Calorimeter. The expresaiona for heat content of a Uquid and its vapor, and the 
heat oontent of sup^heated steam, are made use of in finding the part z of a mixture that exists as 
wet vapor, within certain limits. 

This is commonly known as the determination of moisture or " priming " in steam by means 
of the " throttling " W superheating calorimeter and the necessary data applicable to the above 

TABLE 5 

PROPERTIES OF SATURATED EETEAH 













































HwtCantat 


LaUntBwt 11 










an 


Itaik, 


Si: « 

CD.Pt. Y 


r^ 


laB.l.u. 


In B.t.u. 


TJ 










e. 
















la. el 


Sq.C. 






p-Lb. C 


bf 


ol 




In- 


or of 


Vij«r- 


ol 


II«»y 






if- 




UquId 




■ulkn 


Krul L 


Iquid u 


iSai V 




9 


- 


. 


r 


!/•" ; C 


t" 


f 


P 


tf 


/T 


•" 










~d 


« 


B 






li 




■ 


t.ose 




It 


n 


ISSS 


oosoo 


88 76 


064 


0868 


9789 


1827 


8448 


^776 


*.m 




11 




78 


t 


OMTS 


H 


s 






082 


S 


967 








9208 


<.■ 


.MO 






U 




00646 














9S4 








9108 


4.8 


.US 




I* 




8 


006TI 


DO 


4 


IB 




DIS 


s 


968 


S 


1864 


7214 


9068 


■.loa 




u 


19 


IB 


T 


ooau 


09 


8 


21 




018 




^ 




8009 




S8T1 


a.iu 








M 














007 






£199 




8887 


ID. ISO 














SO 




81 






6 


BS3 




8848 




8468 


as- 




n 




Si 




01614 


87 




86 




897 




928 


2 


£478 


6886 


8soa 




» 


U 


SI 


7 


01664 


4« 




87 




998 




K8 




2681 


6608 




M.» 










85 




60 




40 




989 








£67G 


6402 


80TT 


u.n 








41 




0E86S 


66 




48 




988 


8 


»IS 


6 


8769 


6Z2S 


798S 


n.M 






!1 


SB 


43 


02602 


61 








98S 


8 


918 


t 


8886 


£068 


7897 


It.4* 






rE 


U 


M 




66 








980 


6 


909 




2906 




7881 


U.4I 








B3 




^828 






47 




978 




tot 




2989 


478S 


7761 


5S 




« 


9S 


tt 




78 




48 




976 


"e 


908 


2 


8028 


466S 


T687 


w.w 






56 


18 


M 




TJ 




BO 




978 


3 










7688 


».« 


i.WI 






80 




03780 


SO 












898 


B 






7689 


M 


U.74 


^ 


18 


M 


7E 


03TS« 






61 




971 


' 


898 




8128 1 


448S 1 


7687 



32 



POWER PI^NTS AND REFRIGERATION 



TABLE &^(ConHnued) 

PROPERTIES OP SATURATED STEAM 
(G. A, Goodenougk) 



Absolute 


i 






Heat Content 


Latent 


, Heat 




Entropy 




Pres- 
■ure. 


Teinp., 
•P. 


Volume, 
Cu. Pt. 
per Lb. 


Weight, 
fjb. per 
Cu. Ft. 


in B.t.u. 


in B.t.u. 1 


of 




IJb, per 
Sq. In. 


of 


of 


of Vapor- 
ization 


In- 1 


of Vapor- 
isation 


of 










Liquid 


Vapor 


temal i 


Liquid 


Vapor 


P 


1 


f" 


.1/i" 


t'- 


1 


r 


P 


a' 


r/T 


•* 








d 


9 


H 






1 

n 




9 


16 


216.3 


24.76 


0.04088 


184.3 


1153.4 


969.1 


896.8 


0.8184 


1.4837 


1.7G21 


18 


222.4 


22.18 


.04608 


190.6 


1156.7 


965.2 


891.4 


.8274 


1.4158 


1.7427 


20 


228.0 


20.10 


.04976 


196.0 


1167.7 


961.7 


887.8 


.8356 


1.8987 


1.7848 


22 


388.1 
287.8 


18.88 


.05U 


201.2 


1159.6 


968.4 


888.6 


.3430 


1.8887 


1.7287 


24 


16.95 


.0690 


206.0 


1161.3 


956.3 


880.1 


.8499 


1.3698 


1.7197 


26 


242.2 


16.78 


.0686 


210.4 


1162.8 


952.4 


876.8 


.8668 


1.8670 


1.7133 


28 


246.4 


14.67 


.0681 


214.6 


1164.3 


949.7 


878.7 


.3622 


1.8452 


1.7074 


80 


250.8 


13.76 


.0727 


218.6 


1166.7 


947.1 


870.7 


.8679 


1.8840 


1.7019 


82 


254.0 


12.95 


.0772 


222.4 


1166.9 


944.6 


867.9 


.3731 


1.3236 


1.6867 


84 


267.6 


12.24 


.0818 


226.9 


1168. 1 


942.2 


865.2 


.3781 


1.8137 


1.6918 


86 


260.9 


11.60 


.0862 


229.4 


1169.2 


939.9 


862.7 


.3829 


1.8044 


1.6878 


38 


264.2 


11.03 


.0907 


232.6 


1170.3 


937.7 


860.2 


.8874 


1.2956 


1.6880 


40 


267.2 


10.61 


.0961 


235.8 


1171.8 


936.6 


857.8 


.3917 


1.2871 


1.6788 


42 


270.2 


10.04 


.0996 


238.8 


1172.2 


938.6 


855.6 


.8968 


1.2791 


1.6749 


44 


278.0 


9.61 


.1040 


241.7 


1178.2 


981.6 


868.3 


.3998 


1.2714 


1.6712 


46 


276.8 


9.22 


.1086 


244.6 


1174.0 


929.6 


861.2 


.4086 


1.2640 


1.6676 


48 


278.4 


8.86 


.1129 


247.2 


1174.8 


927.7 


849.1 


.4072 


1 .2670 


1.6642 


60 


281.0 


8.63 


.1173 


249.8 


1176.6 


925.9 


847.1 


.4108 


1.2501 


1.6609 


62 


288.6 


8.22 


.1217 


252.8 


1176.4 


924.1 


845.1 


.4142 


1.2486 


1.6677 


64 


286.9 


7.93 


.1261 


254.7 


1177.1 


922.4 


843.2 


.4174 


1.2873 


1.6647 


66 


288.2 


7.67 


.1304 


257.1 


1177.8 


920.7 


841.4 


.4206 


1.2811 


1.6517 


68 


290.6 


7.42 


.1348 


259.5 


1178.5 


919.0 


889.5 


.4237 


1.2252 


1.6489 


60 


292.7 


7.18 


.1892 


261.7 


1179.1 


917.4 


837.8 


.4267 


1.2196 


1.6462 


62 


294.9 


6.97 


.1486 


263.9 


1179.7 


915.8 


836.0 


.4296 


1.2189 


1.6485 


64 


296.9 


6.76 


.1479 


266.1 


1180.8 


914.8 


834.3 


.4324 


1.2086 


1.6409 


66 


299.0 


6.67 


.1622 


268.2 


1180.9 


912.7 


832.7 


.4352 


1.2032 


1.6884 


68 


801.0 


6.89 


.1666 


270.2 


1181.5 


911.2 


831.1 


.4379 


1.1981 


1.6860 


70 


802.9 


6.22 


.1609 


272.2 


1182.0 


909.8 


829.5 


.4405 


1.1931 


1.6R86 


72 


804.8 


6.06 


.1652 


274.2 


1182.5 


908.3 


827.9 


.4481 


1.1883 


1.6818 


74 


806.7 


6.90 


.1696 


276.1 


1188.0 


906.9 


826.4 


.4456 


1.1835 


1.6891 


76 


808.6 


6.76 


.1738 


278.0 


1183.5 


905.5 


824.9 


.4480 


1.1789 


1.6269 


78 


810.8 


6.61 


.1781 


279.8 


1184.0 


904.2 


828.4 


.4504 


1.1744 


1.6248 


80 


812.0 


6.48 


.1824 


281.6 


1184.4 


902.8 


821.9 


.4627 


1.1700 


1.6227 


82 


813.7 


6.35 


.1868 


283.4 


1184.9 


901.6 


820.5 


.4650 


1.1657 


1.6207 


84 


816.4 


6.23 


.1910 


286.1 


1186.8 


900.2 


819.1 


.4672 


1.1615 


1.6187 


86 


817.1 


6.12 


.1953 


286.8 


1186.7 


898.9 


817.7 


.4594 


1.1674 


1.6168 


88 


818.7 


6.01 


.1996 


288.5 


1186.1 


897.7 


816.8 


.4615 


1.1534 


1.6149 


90 


320.8 


4.906 


.2039 


290.1 


1186^6 


896.4 


816.0 


.4636 


1.1495 


1.6181 


92 


321.8 


4.805 


.2081 


291.7 


1186.9 


895.2 


813.7 


.4657 


1.1456 


1.61ia 


94 


823.8 


4.709 


.2124 


293.8 


1187.8 


894.0 


|12.4 
811.1 


.4677 


1.1419 


1.6096 


96 


824.8 


4.617 


.2166 


294.8 


1187.7 


892.8 


.4697 


1.1381 


1.6079 


98 


326.8 


4.628 


.2209 


296.4 


1188.0 


891.6 


809.8 


.4717 


1.1845 


1.6068 


100 


827.8 


4.442 


.2251 


297.9 


1188.4 


890.5 


808.6 


.4736 


1.1309 


1.6046 


102 


329.2 


4.869 


.2294 


299.4 


1188.7 


889.8 


807.4 


.4755 


1.1274 


1.6028 


104 


880.7 


4.279 


.2337 


800.9 


1189.0 


888.2 


806.1 


.4773 


1.1239 


1.6012 


106 


332.0 


4.202 


.2380 


302.3 


1189.4 


887.1 


804.9 


.4791 


1.1205 


1.6896 


108 


333.4 


4.128 


.2422 


808.7 


1189.7 


885.9 


803.8 


.4809 


1.1172 


1.6981 


110 


334.8 


4.057 


.2465 


805.1 


1190.0 


884.8 


802.6 


.4827 


1.1138 


1.6965 


112 


836.1 


8.988 


.2508 


306.5 


1190.3 


883.7 


801.4 


.4844 


1.1106 


1.6960 


114 


337.4 


3.921 


.2550 


807.9 


1190.6 


8«2.7 


800.8 


.4861 


1.1074 


1.6885 


116 


388.7 


3.867 


.2593 


309.2 


1190.8 


881.6 


799.2 


.4878 


1.1043 


1.6921 
1.6907 


118 


340.0 


3.796 


.2635 


810.6 


1191.1 


880.6 


798.0 


.4895 


1.1012 


120 


341.3 


8.735 


• .26''8 


311.9 


1191.4 


879.6 


796.9 


.4911 


1.0982 


1.6898 


122 


342.6 


8.676 


.2^20 


813.2 


1191.6 


878.5 


796.8 


.4927 


1.0952 


1.5879 


124 


343.7 


8.620 


.2''62 


314.4 


1191.9 


877.6 


794.8 


.4943 


1.0922 


1.5865 


126 


346.0 


8.566 


.2805 


315.7 


1192.1 


876.4 


793.7 


.4958 


1.0894 


1.6862 


128 


346.2 


3.513 


.2847 


316.9 


1192.4 


875.4 


792.6 


.4974 


1.0865 


1.6888 


130 


847.4 


3.461 


.2889 


318.2 


1192.6 


874.4 


791.6 


.4989 


1.0836 


1.6826 


132 


848.6 


3.412 


.2931 


819.4 


1192.9 


878.6 


790.5 


.5004 


1.0808 


1.5812 


184 


849.7 


3.863 


.2978 


320.6 


1193.1 


872.6 


789.5 


.5019 


1.0781 


1.6800 


136 


360.8 


3.316 


.8016 


321.8 


1193.3 


871.6 


788.5 


.5083 


1.0754 


1.5787 


188 


352.0 


3.270 


.3068 


823.0 


1193.5 


870.5 


787.4 


.6048 


1.0727 


1.6776 


140 


863.1 


8.226 


.8100 


324.2 


1193.7 


869.6 


786.4 


.5062 


1.0700 


1.5762 


142 


364.2 


3.182 


.8142 


325.3 


1193.9 


868.6 


785.4 


.5076 


1.0674 


1.5760 


144 


366.8 


8.140 


.8184 


826.6 


1194.1 


867.7 

1 


784.5 


.6090 


1.0648 


1.6788 

• 



WATER, STEAM, AND AIR 
TABLE 5— (CwUinwaO 



POWER PLANTS AND REFRIGERATION 



Suparhaat. Deg. F. 
na, M. Mbak Specific Bkat CuKna. 
(C. A. C 



WATER, STEAM, AND AIR 



35 



If the steam is absolutely dry and saturated in the main steam pipe the total heat //is 1106.0 
B.t.u. per pound, and at atmospheric pressure the total heat H per pound of dry saturated steam is 
1151.7 B.t.u. As the heat content must be the same after free expansion as before there is avail- 
able 1196.9 — 1151.7 or 45.2 B.t.u., which goes to superheat the steam at the lower pressure. The 

amount of superheat or the number of degrees above the saturation temperature, corresponding to 

i 

45.2 
atmoephiCTio preeBure, to which the steam after free expansion will be raised is -— » 96.2^, where 0.47 

0.47 

is the mean q>ecific heat of superheated steam at atmospheric pressure. Henpe the lower thermometer 
will read 212 + 96.2 « 308.2^ F., if no moisture is present in the original steam. 




To Atmosphere 



Pig. 15. Trbottuno OxLOBncETER and Sampung Nozzle. 



If the original steam contains, say 1 per cent of moisture, it will take B.5 B.t.u. to evaporate 
this moisture at 370.7® F. since the latent heat at this temperature is 854.2 B.t.u. per lb. We will then 

36.7 
liave left for superheating 45.2 — 8.5 = 36.7 B.t.u. or the steam will be superheated only 



0.47 



78.1" F. 



It 18 readily seen that as the moisture increases less and less heat will be available for 
superheating, until finally no superheating will occur and the limil of moisture determination by 
the throUHng calorimeter for steam at this pressure will have been reached. 

Tlie general formula for finding the quality of steam by this apparatus at any pressure is 
given below: 

Hjg = Hg where H^ — x n + gi « total heat of one pound of steam at the initial 



Hg »= r J + ga + Cp (tg — tt) =* total heat of one pound of steam at the final or atmos- 
pheno pt'euuure. 

Hence x n + gi ^rt + qt + Cp (t, — tt) 



X = 



ri 4- 9j + Cp {tg - ft) - gi 



36 POWER PIjVNTS and RKFRICIRRATION 

n ood ri = latent heat of vaporiuitioD &t the initiul and final prenurea le^MctJvely. 
Qi and gi ■= heat <A lh« liquid at the initial and final pressures respectively. 
Cp - mean BfK^ilic heat of miperheatcd uteam (see Fig. H). 
t, » temperature o( uteam after supcrhenting. 

It <■ temperature of saturated steam at the final preasure (atmos>bcTe). 
TTie limit of moisture or maximum value of 1 — a;, ia found by making (, = li for any given 
COM and solving f<a- z. These limits range from 2.88 per cent at 50 lb. gage to 7.17 per omt 
moisture at 250 lb. gage, at sea level. 

Practically there are slight errorB in the proceei due to the exposed stem <rf the tbarnometer, 
and the radiation bies From the instrument, The stem oorrection can be made as already indi- 



Via. la CoiiiPACt TBRonuHf} Fio. 17. Sepaiu 



cated, and by heavily lagging the instrument the radiation loss can be largely overoome. Botli 
erttna tend to reduce the reading of the lower thermometer, (,. 

A very compact form of the throttling calorimeter is shown in Pig. 16. 

F<H' VOT "wet" steam a geparatino eaionmeter must be used, and a section <A such an ap- 
paratus is shown in Fig. 17. This apparatus is in effect a small separat^M- which mechanically 
separates the entrained wat«- from the steaoi and collects it in a reservmr (R) where its amount 



WATER, STEAM, AND AIR 37 

IB indicated in a gage glass (G), while dry steam only escapes at the orifice (O). This orifice is of 
known size, and if the pressure in the chamber (C) is known the weight of dry steam passing the 
orifice can be calculated, or a gage (P) can be calibrated to read directly, the weight of steam 
ftowing in pounds, provided the absolute pressure is not less than 25.37 lb. where the orifice dis- 
charges into the atmosphere. For absolute pressures lower than this a calculation must be made 
as stated by the f(H*mula under '' Flow of Steam through Orifices/' 

Mixtures of Air and Satorated Water Vapor. The method of calculating the weight of 
1^'ater vapor mixed with air, for various conditions of pressure and temperature, will be found in 
the Chapter on " Cooling Ponds and Towers.'' A table and diagram are included for convenience 
in solving problems relative to the subject. 

Vlow of Steam Througb Pipes. Various formulas for the flow of steam throu^ pipes have 
been advanced, having their basis upon BemouUi*8 theorem of the flow of water through circular 
pipes with the proper modifications made for the variation in constants between steam and 
water. Unmn's formula based on WeiiAach*8 work is very commonly used and may be stated us 
fotkms: 

2 L 1^ 

*«/X-7rX— See "Friction Head due to Fk>w of Water" (I) 

D g 

in which h r^resents the loss of head in feet of the fluid flowing, in this case steam, which is 
psasing with a velocity of v feet per second, through a pipe D feet in diameter, and L feet long; 
g represents the accderation due' to gravity, and / the coefficient of friction. 

Numerous values have been given for this coefficient of friction, /, which, from experiment, 
apparently varies with both the diameter of pipe and the velocity of the passing steam. There 
are no aatiientic data on the rate of this variation with velocity, and, as in all experiments, the 
effect of change of velocity has seemed less than the unavoidable errors of observation, the co- 
efficient is assumed to vary only with the size of the pipe. 

Unwin established a rdation for this coefficient for steam at a velocity of 100 feet per second. 



^-^O+i^) 



(2) 



where K b a constant experimentally determined, and D the iutcnml diaiiietcr of the pipe in feet. 
If d represents the density of the steam or weight per cubic foot, and p the loss of pressure 
due to friction in pounds per square inch, then 

hd 
p-tt: (3) 



144 
and from equations (1), (2), and (3), 

72 g 



P = r-|xic(i+^J (^> 



To convert the velocity term into weight and to reduce to units ordinarily used let Di » the 
diameter of pipe in inches » 12 D, and w ~ the weight of steam in pounds per minute; then 






w 

4 

9.6 w 



and, V = 



Substituting this value and that of D in formula (4). 



. = 0.04839 /f (if-)— (5) 



38 



POWER PLANTS AND REFRIGERATION 



Some of the experimental determinations for the value of K for steam are: 

K - 0.0026 {ft. C. Carpenter), 
K - 0.0027 ((7. H, Babcock). 

Substituting the value 0.0027 in formula (5) gives, 

/ 3.6\ tr* L 
^ = 0.000131 (l + -) X — (6) 



and, w = 87.5 



V d Z>,» 






(7) 



in which the various symbols have already been defined.* 

This formula is the one most generally accepted in this country for the flow of steam in pipes. 

Equation (4) may be written, 



V « 16,060 
Equation (6) may be written, 



Ld 



(■ - 1!) 



, in which V — velocity of 
the steam in ft. per min. 



p ^ A X 



w^L 



(8) 



0.000131 I 1 + 



in which A = 

Equation (7) may be written, 



f. 3.6\ 



w 



u; = C 



K]* 



(9) 



in which 



C= 87.5 



D,» 



(■ ^ -»•) 



i 



For values of A and C see Table 6. 



Equivalent Length of Pipe for Each Globe Valve, Entrance, and Elbow. In addition to the 
loss of pressure due to friction, in straight pipe, there is also a loss of pressure due to a change in 
the velocity of the steam at the entrance to the pipe. This drop in pressure due to getting up 
velocity in the pipe is very slight and is seldom taken into account. 

Elbows, globe valves, and a square-ended entran e to the pipe, su h as occurs when steam 
is taken off through a tee at right angles to the main, all offer resistance to the flow of steam, 
thus causing a drop in press re, which should be taken into account and proper allowance 
made for it. 

Friction is greater through s'lort radius elbows and tees, than through elbows and 
tees of long radius. The resis ance offered by a globe valve is about H greater than that due 
to a short radius elbow, whereas gate valves offer practically no resistanc ^ to the flow, providing 
they are opened wide. The resistance offered by a square-ended opening, or at the outlet of a 



*d^ the density, is taken as the mean denai^ at the initial and final 
up to 6 diametera actual internal diameters should be used. 



pressures and in exact work on pipe* 



WATER, STEAM, AND AIR 39 

tee where a branch ia taken <^ at ri^t angles, is about the same as that for a globe valve having 
the aame si« opening. The resistaoce offered by a loi% radius pipe bend is very slight and 
may be t&kea aa equal to the reaistasce tiered bjr the same length of straight pipe, or in other 
words, all pipe bends niay be considered as strai^t pipe of equal lei^h. 

TABLE e 



SSSSLSff 


iKtm-ot 


VmliHOl 
Comttnt-C" 


V*|UM Of CoMtant "A" 


b« Added 


to^ Added 




MT 


■1 

MlO 
2,01S 
2.TSE 

irm 

11 

ISSiSTO 




J 

i 

SB 

1 

43 

200 


1 6 






1 
1 










I 


■- 






































I 

a 









































It ia cuBtomai? to CMinder the reeiBtance offered by valves and fittings, etc., as equivalent 
to a length of strai^t pipe which will offer the same resistance, or cause the same drop in pres- 
Rure. When this equivalent length has been determined it should be added to length L in the 
formula, and p, or vt, computed accordingly. 

Equivalent length of straight pipe, in inches, to be added for each globe valve, oT square- 

ended opcnmg = se;- 

(^ + "a) 
Equivalent length of straight pipe, in inches, to be added for each 9(MIeg. elbow in the line 

C + MY 

Where Di = inside diameter of pipe in inches. The values in Table 6 have been com- 
puted from the above formulas. 

Bzanpla: Let it be required to determioc the pt«Bsure loss in a pipe line for the foUowiDs con- 

Di = 5" L " aoC u - 250 lb. 
9te»m pressure •• 150 lb. sage, or 105 lb. iU>wdute. 



- - 0.363 (troni Table 5). 



■ Notx. — ^AH pipe 14 inebae dl 



40 POWER PLANl^ AND REHIIGERATION 

Ftem Table 6, value of constant A « 0.000.000,07. By substitution in equation (8) 

250* X 300 

p - 0.000.000.07 X - 3.62 lb. per sq. in. 

0.363 

Steam Flow Chart The use of steam flow charU based on the above fonnuias is very general 
in engineering practice, and a variety of these charts have been prepared using various coordinates 
depending on the relations which are to be expressed. Thus charts may be laid out to show 
velocity of flow, weight of steam, or pressure loss. The latter value is most often required in 
propCHtioning a piping system, and the following logarithmic chart, Fig. 18, by PrqfeMor H. V. 
Carpenter will be found very useful, as it shows the relatiim between size of pipe, average pressure, 
drop in pressure, and weight of steam passing in pounds per minute. 

Examples. Follow the heavy dotted lines, and assume an allowable pressure loss of 0.3 lb. per 100 
ft. for a 3-in. pipe at an average pressure of 80 lb. absolute. The weight of steam delivered will be 21 
lb. per min. Again, assume a drop of 1 lb. per 100 ft. for a 10-in. pipe ddivering 860 lb. xier min. 
The average absolute pressure must be 60 lb. per sq. in. Finally, assume a 20-in. pipe is delivrainn 
4,000 lb. per min. at an average absolute pressure of 250 lb. per sq. in. The drop in pressure will be 
0.15 lb. per 100 ft. of pipe. 

Professor Carpenter says, regarding the accuracy of the charts: " They represent the for- 
mulas exactly, except for the inaccuracies in drawing and in reading the scales. These errors are 
far within the limits of accuracy needed in practice so the charts may be used with the same 
degree of confidence as the formulas. 

'' As to the accuracy and range of the formulas, it seems that all the published experimenta 
were made with pipes of from 1.85 to 4.0 in. in diameter. There is little doubt that the f<Hmulas 
may be applied with entire safety over a much wider range than this, but the practical limits 
are unknown." 

Flow of Steam Through Orifices. The flow of steam from a higher to a lower pressure in- 
creases as the difference in pressure increases to a point where the absolute terminal pressure he- 
comes 58 per cent of the absolute initial pressure. Below this point the flow is not increased by a 
reduction of the terminal pressure, even to the extent of a perfect vacuum. The lowest initial 
pressure for which this statement holds, when steam is discharged into the atmosphere, is 25.37 
lb. For any pressure below this figure, the atmospheric pressure, 14.7 lb., is greater than 58 per 
cent of the initial pressure. 

Napier deduced the following approximate f<»*mula for the flow of steam through an orifice. 

va 
70 

Where W = the pounds of steam flowing per second, 

p B the absolute pressure in poimds per square inch, 
and a = area of the orifice in square inches. 
In some experiments made by Professor C. H, Peabody on the flow of steam through pipes 
from }i in. to 1 ^ in. long and ^ in. in diameter, with rounded entrances, the greatest difference 
from Napier's formula was 3.2 per cent excess of the experimental over the calculated results. 

For steam flowing through an orifice from a higher to a lower pressure where the lower pres- 
sure is greater than 58 per cent of the higher, the flow per minute may be calculated from the 
formula: 



Where W » the weight of steam discharged in pounds per minute, 
A a area of orifice in square inches, 
P » the absolute initial pressure in pounds per square inch, 



WATER, StEAM, AND AIR 
Arerage Steam PreMore, Pounds per Square Inch Abeolute 



41 




All ai5 

ViO. 18 



q;9 



10 



04) OJ OJi 0.6 0^ 1 1.5 8 3 4 

Lonof Pressure per 100 Peet of Pipe, Pounds per Square Inch 

Obabt Sbowino Loss of Prbssubb when a Givbn Amount of Steam Per Minute 
IS Deuybked ihrouob a Pipe op Givbn Size. — H. V. Carpenter, 



42 POWER PLANTS AND REFRIGERATION 

d — the difference in prewure between the two sides in pounds per square iiuli, 
K •• A coDatant — 0.93 for ^ short pipe, and 0.63 for a. hole in a thin plate c»' a 
safety valve. 
BzamplB. Let it be required to determine the weight of steam flowing per min. from a boiUr Into 
tbe atmosphere through a short length of 1-in. pipe, for the rollowios ci 
Initial presaure in boiler (p) — 100 lb. absolute. 
Internal area of 1-in. standard pipe <a) - 0,804 sq. iu. 
By subatitulioo in Napitr'a formula 



70 



weight per tnin. ° 60 X I.20S - 72.48 lb. 



H«flsurem«0t of Steam Flow. All steam meters for either indicating or recwding iba 
weight at steam flowii^ in a pipe are based on the following law: 

W ^ AdV 
in which 

W ■■ weight of steam flowing per sec. 
A s iiit«mal area of pipe, sq. ft. 
d = density of steam. 
V = velocity, ft. per sec. 
The density of steam is a function of the pressure and the quality, x, if it ia wet saturated 
which is the usual condition in practice. The quality may be determined by means of a throt- 
tling or separating calorimeter previously described. The velocity in the Pilot tube type of 
meters, of which the Gtneral EUctrie Co.'t and the Gebhardt types are examples, is determined 



Fio. 10. Phincipu: of tsb Pitot-Tubii Ttpb or Stiaii MnvB. 

fnMU the velocity head or piesBure, measured by the height of a colunw of water or mereuiT 
supported by this head or pressure (Fig. 19). 

The static head or pressure on the liquid column W is tnuumitted through the upper con- 
nection s while the total or dynaroic pressure is transmitted to the liquid column by means of 
the tube D bent at right angles to the 6ow. 

Hie height H of the liquid column is a measure of the dUTerence between the total and 
static pressure, and is therefore an indication of the velocity head or pressure existing at the point 
ot measurement. The relation between the height of the liquid in the tube and the velocity 
at the center of the pipe is determmed from the following equation: 

.... K = C •f2^h (1) 

in which 

V ^ mean velocity of flow over entire cross section, ft. per sec. 
A = height of a column in feet of the medium flowing. 



WATER, STEAM, AND AIR 43 

C B a coefficient to correct for the average rate of flow as determined by experiment for 
various sizes of pipes. 
Tbe actual measurement of the velocity head is made in inches of water or mercury. 
Let k — density of the liquid used in the tube. 

d » density of the steam. 

U « velocity head measured in inches of the liquid used in the tube or manometer. 

12 dh=^kH or h^- -— -. 

12 d 

Substituting the value of A in (1) 



^ 



gkH 

-IT ^^^ 



The commercial form of this type of meter gives results within 2 per cent of actual condenser 
weights for velocity pressures corresponding to 1 inch or more of water. 

Tbe calibration of the indicating colunm to read the weight of steam flow direct is beet 
made by weighing the water from a condenser to which the steam is delivered. 

For a description of various forms of steam flow meters see Carpenter and Diedericha ''Ex- 
perimental Engineering,'' also " Steam Power Plant Engineering '' by G. F. Gebhardt, 

AIR AND OTHER GASES 

Prop er t i es of Air and Other Oases. Air is the most general example of a so-called perfect 
or permanent gas to be found in nature, |md like the other so-called perfect gases conforms more 
or less closely to the laws of perfect gases. These laws are stated in the following paragraphs. 

Pure dry air is a mechanical mixture of oxygen and nitrogen, that is, the oxygen and nitrogen 
can be separated from each other by purely physical means. This mixture is made up as follows: 

By Volume By Weight 

Oxygen 20.91% 23.15% 

Nitrogen 79.09 76.85 

Air as found in nature alwajrs contains other constituents in varying amounts such as carbon 
dioxide, ozone, water vapor, dust, bacteria, etc. SeetheChapter on *'Ventilationand Air Analysis."* 

The specific dentily, or weight per cu. ft. of dry air decreases with the temperature, and 
conversely the specific volume, or volume per pound, which is always the reciprocal of the density, 
increases with the temperature. See Table 7 for properties of dry air. 

The specific heat of air at constant pressure, or the B.t.u. required to raise one pound 1° F. 
at the pressure of the atmosphere, varies from 0.2375 to 0.2430 as determined by various investi- 
gators. The value 0.24 is recommended for engineering calculations. 

It has been found that a given volume of air expands when heated under constant pressure, 
and again that if the temperature of a given volume of air is kept constant and the pressure in- 
creased, contraction takes place. These changes follow perfectly definite laws, w^hich apply to 
other gases as well as air, known as '' The Laws of Perfect Gases." These laws do not apply 
to steam, since it is not a perfect gas. 

Bcyle^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 in- 
versely as its absolute pressure. Hence if Pi and Pt represent the initial and final absolute pres- 
sures and Vi and Vt represent corresponding volumes of the same mass, say 1 lb. of gas, then 

V P 

— ■■ -z-or PiVi « PiVt, but since PiVi for any given case is a definite constant quantity, it 

Vt Pi 

*VoIiiiimL 



44 POWER PLANTS AND REFRIGERATION 

follows that the product of the absolute pressure and volune of a gas is a constant, or PF « C, 
when T is kept constant. 

Any change in the pressure and volume of a gas at constant temperature, as indicated above, 
is called an isothermal change. 

Charles* Law refers to the relation between pressure, volume, and temperature of a gas and 
may be stated as follows: The volume of a given weight of gas varies directly as the absohite 
temperature at constant pressure, and the pressure varies directly as the absolute temperature 
at constant volimie. Hence, when heat is added at constant volume F^, we have the equation: 

— = — or for the same temperature range, at constant pressure Pf, the relation ^ "TT —'^' 
Pi i I Vi Tt 

In general we have for any weight of gas A/, since volume is proportional to weight at aiqf 
given volume and temperature, the relation 

PV^ MRT 

which is the characteristic equation for a perfect gas. In this formula 

P » the absolute pressure of the gas in pounds per square foot. 

F » the volume of the weight M in cubic feet. 

M ^ the weight in pounds of the gas taken. 

A " a constant depending on the nature of the gas. 

T = the absolute temperature in degrees F. 

A perfect gas conf(»*ms exactly to the above equation, and while no gases are "perfect" in 
.this sense they conform so nearly that the above equation will apply to most engineering 
computations. 

Another form of the characteristic equation is sometimes used, in which M and R are elimi- 
nated. Let Po» Fo, and Tq denote the initial condition of a given quantity of a gas which undei^ 
goes a change in pressure, volume and temperature, the second condition being denoted by P, F, 

Po Fo 

and T. For the initial condition then P© Fo = MRTq or — — — =* MRj and for the second 

i 

PV 
condition PV. == MRT or — - - MR so that the left hand members of the two equations are 

^ ^ PoFo PF 

equal to each other, or - -~ — = —zr' 

So long as the same units are used for pressure, as poimds or ounces, and the same units are 
used for volume, as cu. ft. or cu. meters, and the temperatures are expressed in the same absohite 
scale it makes no difference what these units may be and the above equation hc^ds. 

In order to determine the value of R for any gas we must know the absolute pressure and 
temperature, and the volume in cu. ft. of one pound. For air at sea level, the absolute pressure 
is 14.7 lb. per sq. in or 2146.3 lb. per sq. ft. and at a temperature of 32^ F. the absolute tempera- 

PV 
ture is 32 + 459.6 = 491.6 **F., and the vohime is 12.39 cu. ft. per 1 lb. Now since R = —=- 

„ 2146.3 X 12.39 _ „^ ^ . 

we have R = ■ = 53.37 a constant for air. 

492 

It follows then that the volume of 1 lb. of air (known as the specific volume) at any tem- 

perature and pressure, can be found at once by the equation F = — —- , and the vahie of 

R for other gases will be directly propcMtional to the specific volumes of such gases and air. See 
Table 8. 



WATER, STEAM, AND AIR 



45 



ISL 



TABLE 7 

PROPERTIES OF DRY AIR 
BarofBetrie Preasure 29.921 Inehet 



lai 

r 

r. 

I 

.7 



Fabr. 



Wdfht Mr Cubie 





5 

10 

15 

20 

25 

80 

S5 

40 

45 

60 

55 

00 

65 

70 

75 

80 

85 

90 

95 

100 

105 

110 

115 

120 

125 

180 

185 

140 

145 

150 

160 

170 

180 

190 

900 



240 



280 
800 
850 
400 
450 
500 
550 



0.08686 
.08544 
.08458 
.08868 
.06276 
.08190 
.08107 
.08025 
.07945 
.07866 
.07788 
.07718 
.07640 
.07667 
.07495 
.07424 
.07856 
.07289 
.07222 
.07157 
.07098 
.07080 
. vd9«o 
.06908 
.06848 
.06790 
.06782 
.06675 
.06620 
.06565 
.06510 
.06406 
.06804 
.06205 
.06110 
.06018 
.05840 
.05678 
.05516 
.05867 
.06225 
.04903 
.04618 
.04864 
.04188 
.08982 
.03746 



Per Cent of Volume 
at 70« F. 



700 


.08428 


800 


.08151 


900 


.02920 


1000 


.02720 


1200 


.02892 



B.t.u. Aboorbed by 

One CubicFoot Diy 

Air per Degfoe F* 



Cubie Foot Dry Air 

Wanned One Degree 

per B.t.u. 



0.8680 


0.02080 


48.08 


.8772 


.02060 


48.55 


.8867 


.02089 


49.05 


.8962 


.02018 


49.56 


.9057 


.01998 


50.05 


.9152 


.01977 


50.58 


.9246 


.01957 


51.10 


.9840 


.01988 


51.60 


.9434 


.01919 


52.11 


.9530 


.01900 


52.64 


.9624 


.01881 


68.17 


.9718 


.01863 


53.68 


.9811 


.01846 


54.18 


.9905 


.01829 


54.68 


1.0000 


.01812 


65.19 


1.0095 


.01796 


65.72 


1.0190 


.01779 


56.21 


1.0288 


.01763 


56.72 


1.0380 


.01747 


57.25 


1.0472 


.01782 


57.74 


1.0570 


.01716 


58.28 


1.0660 


.01702 


58.76 


1.0756 


.01687 


59.28 


1.0850 


.01673 


59.78 


1.0945 


.01659 


60.28 


1.1040 


.01645 


60.79 


1.1183 


.01631 


61.82 


1.1280 


.01618 


61.81 


1.1820 


.01605 


62.81 


1.1417 


.01592 


62.82 


1.1512 


.01578 


68.87 


1 . 1700 


.01554 


64.85 


1.1890 


.01680 


65.86 


1.2080 


.01606 


66.40 


1.2270 


.01484 


67.40 


1.2455 


.01462 


68.41 


1.2833 


.01419 


70.48 


1.8212 


.01880 


72.46 


1.8590 


.01843 


74.46 


1.8967 


.01308 


76.46 


1.4845 


.01274 


78.60 


1.5288 


.01197 


83.65 


1.6230 


.01130 


88.60 


1.7177 


.01070 


93.46 


1.8113 


.01018 


98.24 


1.9060 


.00967 


108.42 


2.0010 


.00923 


108.85 


2.1900 


.00847 


118.07 


2.8785 


.00782 


127.88 


2.5670 


.00728 


137.37 


2.7560 


.00680 


147.07 


3.1655 


.00603 


165.83 



Specific Heat of Gases. Reference has already been made to the fact that gases have two 
speetfic heais, one is the sp^fic heal at constant pressure Cp and the other the specific heal al 
canMant volume Cp. 

The vahie of C, can be found experimentally if we take one pound of gas occupying a fixed 

P V 
vohime Vi at pressure Pi. The absolute temperature is then Ti « . 

R 



Now add heat to this 



gas and its temperature and pressure will become Ps and Tt, No external work has been done 
as the vohime rranained constant and hence all the heat supplied has been uped to raise the tenn 
perature of the gas. See Fig. 20. If H represents the heat added then H = C^ (Tt — Ti) or 



of the SOS I' F. 



POWER PLANTS AND BEFEUGERATION 
, where C, - Bpecific heat at conaUut volume — B.t.u. requited to niaa 1 lb. 



tabu: 8 

THERUAL PROPERTIES OF OASES 



NuMofGu 


•ff 


2« 


Dendty 


Gm 


¥ 


SpwiOc HaM 


*i 

Atmo. 


'^ 


O 


C 


» 




» 


4 


5 


6 


T 


a 






Mas 


nun 


ES " 


in 


ain 


,,, 




A. 
f 

CHt 

1 

aoi 


1 

i 

s 


08 

L 




If. 

1227 

ri 

IWGe2 

E 

08»2 


i 
1 


P 
Is 




1! 
1 

40 

g 

40 
40 
2g 


! 


210 

iS 

ZEO 
420 

1 

461 


' 










OM 








ESii- ::::;;:::: 


m 



























































• FropenlH of than (u*a <iwy traatly with tba tampentun and pmnire. 

The value oF C^ can also be found in aBomewhatsiinilarmanDerif weaasiune wehave 1 lb. of 
gas io a cylinder fitted with a frictionleaa piston which is 1 sq. (t. in area. Initial condition is 
Pi, Vi, and Ti, whete Pi is the couatant weight of the piston. Now odd heat and change the 



^5 I 

For Laws of Perfect Gases 



^7 



volume and temperature to Vi and Tt, but Pi = Pi. In this case we have performed external 
work by raising the piston, as well as increased the temperature oT the gas. The work done 
is equal to Pi {V, — Vi) and its heat equivalent is found by dividing by 778. See Fig. 21. 

If H lepresents the heat added then H = C, {T, ■ 

temperature + work done. 

But it C, = specific heat at constant pressure then C, (T", - T,) - C, (T, - T,) + 

■ ^'^'_^^' — ' nnce Pi - P» Abo P, V, - B T", and P, V. - « T, so that C, (7*, - T,) - 



WATER, STEAM, AND AIR 47 

C^ (Tt — Ti) H -— or Cp ^ Cp -h -— for the gas in questicm. Prom this relation it 

77o 77o 

will be seen that the specific heat at constant pressure is always greater than that at constant 
volume. See Table 8 for values of specific heats. 

EiqMuision and Compression of Perfect Gases. The heat required to change the volume of a 
gas, the relation between the pressure and volume being expressed by some law, such as P V** » 
K (a constant) is found in the following manner. Referring to Fig. 22, it is apparent Pi Vi* » 
K and Pi Ft" = K from which 

— 1 = 1 — I orn=» , so thai; K can be readily found. 

Now the total heat required will be that necessary to change the temperature C, (7s — Ti) 
and do the external work W represented by the area abed. 

This area abed is equal to the smnmation of the elementary areas PdV = I Pd V, but 

^ ^ . ^ nKdV K r -1 -\ K r I 1"!.. 

'^ *= t;s so that W ^ I — T=r r I ^w— ^ == r I tt^tt — t^ »~i • Now sub- 

stitute the value of iC = Pi V*i » Pi V^t in the last expression and we have W 



n- 1 

1 R 

[Pi Vi - P, y,l ft.-lb. or PT = IRTi- R Tt] = :; [Tt - Til and expressed in heat 

n — 1 1 — n 

R r 1 r R ^ 

units = ~— [Tt - Ti]. Hence the heat required is H = C„ + ——- \(Tt-Ti). 

778 (1 — n) L 778 (1 — n) J 

Value of the exponent n. If expansion or contrac tion t a kes place wit hout loss or gain of heat 
the change is said to be adiabatic. In this case no heat is added and hence H ^ =^ ( C, + 

"r^" V ' R 

(Ti " Ti), But as already stated (Cp — C,) = — - and by substitution C„ + 



778(1-11)/ ^ • *' " ' " " 778 

— -^ 0, or Cy — n C, + Ca — C» = 0. From which n = -^ and hence the value of the 

1 — n C, 

exponent for adiabatic compression or expansion of a gas is equal to the ratio of the 
tapea&c heats. 

If we compress a gas adiabatically the work of compression expressed in heat units is equal to 

the heat required to change the temperature. As already stated W » (Pi Vi — Pj Fj), the 

— — ^— ^— — ^— — n — 1 

woric of compression in ft.-lb. But for an adiabatic change ff = = C, (Tj — Ti) -f PT -— 

778 

W 
from which it i^pears that zzr = C, (T'l — Tj). 

778 

Furthermore when a gas is expanded adiabatically the work performed by the gas expressed 

in heat units is equal to the heat abstracted in lowering its temperature. 

The relation between pressure, volume and temperature, for adiabatic compression or ex- 

Q 

panaon, can be expressed as follows, the value of n being -~, and the initial and final states 
being Pi, Vi, Ti, and Pj, Vj, Tj. The characteristic equation of a perfect gas where Af is 1 lb. 

can be stated as i ■■ , and hence — =* *= I — I X i — i . Also, we have smce 

, fi' r. p.F. VpJ \v»J' 



POWER PLANTS AND REFRIGERATION 



■(fr- 



' m' 



Aud therefore 



T, fv,\'-' r, IP,\l 
T,-\.rJ "^v,-\pj" 



Those last three equations may be readily solved by the use <rf a table <rf log^thms. 

MeaBurement of Air Flow. There are several methods employed for measuring the quantity 
o! air delivered by a fan, blower or air c(»npi«eBor. The two methoda most oommcmly employed 
in this connection are (1) by means of a circular orifice and (2) by the Pitot tube. The method 
employing the Pitot tube is fiUly described under the chapt«r on " Hot Blast Heating." 

Tlw Oflflce Method. The discharge from the oompreesor or fan is piped to a gauging box 
similar in constniction to the one shown in Fig. 23. The opposite end <k the box is pnmdtd 





Flo. 23. Details o 



with a circular orifice as shown, discharging directly into the air. The static pressure existing 
ivithin the box is measured by means of a U tube, which indicates the difference in preesuie m 
inchcfl of water between the two sides of the orifice. Hie t«mperature of the air passing through 
the gauging box is also recorded as well as the barometric pressure of the air. The discharge 
from the orifice must be free and unobstructed, so that the pressure on the discharge side will 
always be that of the atmosphere. 

The xoeiiflU of air pasBing the orifice per seocnid is then readily determined by substituting 
in the following equatkn. The ooefficient C to be used in this equation has bem determinad 



WATER, STEAM, AND AIR 



49 



by R, J. DwrUy, and may be taken from the curves in Fig. 24 for various sizes of orifices and 
differeooes in head. A complete discussion of this method oi measuring air will be found in 



XK62 



aei 



0.60 



0.59 



0.68 





Coat 


lelMtt 


of Oitt Hsrg§ 1 ir Varh ut Orli 009 ^ 


iacds 
















Poundi 

a, 


par St 

romoti 
«• Boi 


zondmt 

fc Pfea 
9dln P 


uce 

tur&wgi 
^etea 


?99^, 
tnchi 


;xd* 
a 






^ 




^^Ori lea 






^^Vrlfl' 




^-^ 


^ 


■^ 




>2 










^ 


^ 








. — 


■ 




r 






-^ 


^ 


^ 


^ 














2" 










^ 


^ 


^ 




_ 














P 




^ 
















. 






=: 


^ 




5r 


firm 
























































































/ 




1 


1 


Im 


1 
7hes 


4 
f Wat 


er 


5 




5 







Pig. 24. CoBFiTcnBifrs of Discbabob fob Yajuous Orifices and Heads. 



VoL 27 of the ''Transactions of the A. S. M. £/' under "Air Flowing into Atmosphere through 
Circular Orifices." 



W - 0.01369 X C + 



Wt. 



in which 



W » weight of air flowing in pounds per sec. 

C « a coefficient depending on values of d and i (see chart, Pig. 24). 

d « <liameter of orifice in inches. 

i B difference in pressure in inches of water between the two sides of the orifice. 

T » absolute temperature of the air passing the orifice » 460^ + f"' 

t » degrees Fahrenheit in gauging box. 

P B pressure in pounds per sq. ft. of the atmosphere based on barometer reading. 
The above formula may be used fw any atmospheric pressure, but for 30" barometric pres- 
sure the formula reduces to: 



W - 0.6299 



c*^||. 



flow of Air through Pipes and Ducts. The same general formub as used for the flow of 
water may be applied, with sufficient accuracy, for air flowing under low pre89un$ as in 
veotOatiiig dactB» flues and chimnesrs. 

The flow of air under low preesiu-es is fully discussed under the Chapter on " Hot Blast 

Heating." • 
• Vohinie L 



50 



POWER PLANTS AND REFRIGERATION 



s 

g 






o 
o 

o 

^ i 

Q 

B 



s 






CO 



^ 



So 



to 



^ 



00 



lb 



09 



CO 






i 

M 



v^ 



;^ 



8. 



i 



a 
S 

I 
1 

«a 

s 

-a 

•am 

•2 



o 

s 



II 

J 



04 00 f-4 ^ o ^ ee 



• • « 



••^sssssss 



o<0tooo^eOMe»^^ 

"^ i i i u i i i i s 



ooe"<-toook^eoei<-teaoo 

^* s i 8 5ji s ^* s gj I g 

^ ^^ CO ^ 



• ••••••••••• 

y^ <■« c< eo "P ^ 



le o « <-t 
01 00 -^ ei 



v4N^OOC400^01Q9«00 



Okoooo>oo>o too 

« S 2 S* S 2 I g S 1 1 1 1 



to eo «-i Q 



S S2 S 2 S II § g I g g I S 

11 f-< 01 CO lO 



'<'' S ^' S 2 si S' 3* S S ^ I S S S S S 

* ^ 5 • • « S g og 3 g g 

ot-o>oo<e^i-*'*»o^ 

" • d ^' 5 ? S S S I S I g 11 S 2 I S 

•-4 v« 09 



oi -*' « 15 ^ S 8 S S d § § ?2 S S g S g S 2 

11 1-t 04 ^ 



01 



(0000040000000 

^ ^- »o- o o* g 3 5 g g » g o o g o g 



S 



_ - _- »o d t- 

f-l iH rH 09 ^ >0 



~o>«Hi4*aoo9toooooo>eokeo9 

.-« eO ^ t- 09 b- rH 



iH^C9'9oO>eQOON^09 

1-1 iH ^ fco 09 e- j-j 



So o 
«H 09 



„ . . .^ _, e» 00 ;:< 



00O>(OO900C9Ok09090Oa0 

• •••••••••• 

eo to 00 JO 15 35 g ^ ;2 o g 

09 eo 00 



S et «H S o> 

09 W to 



(0O9O>tO0S-<4*iHCoaOOO 



•«*ooo9ooe9;H'jj-jto 

^ rH eo to t- CO o 



^ooooo9<-)-^<-<$SSS9 
r-ieotocoeoo«M 

11 09 ^ 00 



^lOtOO^A t^t^O 

tO^Q09'«>:jaO:f|:09e9QO«MOOOO 

r^Sco-^ooMgoj^HtdJo&ot-i-o 

r^f-ieo>o^40(o^aD9r<'v4 

r-l09^00O9o5|jjH 
,1 »H 00 to 



0»~00 009tOCotO toto 

■ • • • 

eo 00 to to 



^e9eO>dOiOt0090QO 

»H iH CO «o 2 g » 

*-• 09 CO 



s 



lO O 00 1-t '^ to (O 



tOtotoaoeOtiMco«0^;2^2fi 

rHCOV^V^-lAlO'^toiiCO'jMS 

^O9"*»oo<o«e9googca 



0> 09 O to ^ 



QM-^OQlO^tO^tOOpSOO 



r-i eo 



o^wOflO^$9^ 

^^09'<«*l00(0« 

11 ^ eo 



09 O 

to il Q to O to 

09 00 O 09 O 09 

«H 04 >0 ^4 O ^ 

H 09 00 



09 




St-SSoSoo 

n 1-1 01 09 CO <« 



iH .1 09 eo "«* 



•1 f-t 09 09 CO eo ^ 



WATER, STEAM, AND AIR 



51 



TABLE 10 

PRESSURES AND SQUARES OP PRESSURES 





2 

4 

6 

8 

10 

12 

14 

16 

18 

20 



24 
26 
28 
80 
82 
84 
86 
88 
40 
42 
44 
46 
48 
50 
68 
64 




14.7 
16.7 
18.7 
20.7 
22.7 
24.7 
26.7 
28.7 
80.7 
82.7 
84.7 
86.7 
88.7 
40.7 
42.7 
44.7 
46.7 
48.7 
50.7 
52.7 
54.7 
56.7 
58.7 
60.7 
62.7 
64,7 
66.7 
68.7 



oflg 



OQ 



<& 



216 

279 

350 

428 

515 

610 

713 

824 

942 

1069 

1204 

1347 

1498 

1656 

1823 

1998 

2180 

2372 

2570 

2777 

2992 

8215 

8446 

8684 

8981 

4186 

4449 

4720 



56 
58 
60 
62 
64 
66 
68 
70 
72 
74 
76 
78 
80 
82 
84 
86 
88 
90 
92 
94 
96 
98 
100 



70.7 
72.7 
71. 7 
73.7 
78.7 
80.7 



,7 
.7 

.7 
.7 
.7 



82 

84 

88 

88 

90 

92.7 

94.7 

96.7 

98.7 

100.7 

102.7 

104.7 

106.7 

108.7 

110.7 

112.7 

114.7 




4398 

52S5 

5380 

5833 

C194 

6312 

6839 

7174 

7517 

7868 

8226 

8593 

8968 

9851 

9742 

10140 

10547 

10962 

11385 

11816 

12254 

12701 

13156 



105 

no 

115 
120 
125 
130 
135 
140 
145 
150 
155 
160 
165 
170 
175 
180 
185 
190 
195 
200 
210 
220 
230 



119.7 

124.7 

129.7 

134,7 

139.7 

144.7 

H5.7 

154.7 

159.7 

154.7 

1G9.7 

174.7 

179 

184 

189 

194 

199. 

204.7 

209.7 

214.7 

224.7 

234.7 

244.7 



.7 
.7 
.7 
.7 
.7 




14328 
15550 
16822 
18144 
19516 
20938 
22410 
23932 
25504 
27125 
28790 
30500 
82290 
34100 
35980 
37905 
39875 
41900 
43970 
46090 
50490 
55060 
59860 



240 
250 
260 
270 
280 
290 
300 
825 
350 
875 
400 
425 
450 
475 
500 
550 
600 
650 
700 
750 
800 
900 
1000 



.7 
.7 
.7 
.7 



254.7 
264.7 
274.7 
284 
294 
804 
814 
839.7 
364.7 
389.7 
414.7 
439.7 
464.7 
489.7 
514.7 
564.7 
614.7 
684.7 
714.7 
764.7 
814.7 
914.7 
1014.7 




64855 

70055 

75450 

81050 

86845 

92840 

99040 

115400 

182940 

151850 

191950 

198800 

215925 

289790 

264900 

818900 

878900 

441800 

510800 

584800 

668750 

836700 

1029650 



Flow of Compressed Air in Pipes. The variation of density with variation of pressure duo 
to the elasticity of air makes a determination of the friction losses accompanying its passage 
through pipes a more complicated matter tlian the calculation for water-friction losses. Water 
being of practically constant density under all ordinary pressures, its rate of flow through a pipe 
of uniform dian:eter will be imiform throughout the length of that pipe, in spite of the decreas- 
ing pressure accompanying its progress. The friction losses through a unit distance— say 100 
feet — in any part of the pipe line will therefore be the same as the loss through an equal distance 
in any other part of tlie pipe; or, in other words, the losses are directly proportional to the length 
of straight pipe. Air, on the other hand, enters a pipe at a certain pressure and velocity; as 
it advances through the pipe a certain loss of pressure occurs in overcoming frictional resistance; 
this loss of pressure is, however, accompanied by an increase of volume, and a corresponding 
increase in velocity of flow. This variation in velocity of flow throughout the length of the 
line results in a variation in frictional resistance, and the loss of pressure in a unit distance is 
the same at no two points in the pipe. 

Table 9 is based on the formula of J. E. Johnamij Jr., published in the "American Machinist," 
July 27, 1890: 

0.0006 V^ L 
P.^-P.^- ST-' 

in which Pi = absolute initial air pressure, lb. 

Ps = absolute terminal pressure air, lb. 

V = free air equivalent in cubic feet per minute of volume passing through pipe. 
L — length of pipe, feet. 
D « diameter of pipe, inches. 

The " free air equivalent " referred to above is the volume measured at atmospheric pressure. 



CHAPTER III 



FUELS AND COMBUSTIOlf 

FUELS 

Clasiifiaition* Fuels are generally claasified as solid, liquid, and gaseous. 

Solid fuels are coal, wood, and wastes. 
Liquid fuels are petroleum and its products. 
Gaseous fuels are natural and artificial gas. 

SOLID FUELS— CX)AL 

The Formation of Coal. All coab are of vegetable origin, and are the remains of prehistorio 
forests. Destructive distillation, due to great pressures and temperatures, has reserved the 
organic matter into its invariable ultimate constituents, carbon, hydrogen, oxygen, and other 
substances, in var3ring prop(H*tions. The factors of time, depth of beds, disturbance of beds, and 
the intrusion of mineral matter resulting from such disturbances have produced the variation 
in the degree of evolution from vegetable fiber to hard coal This variation is shown chiefly 
in the content of carbon, and Table 1 shows the steps of such variation. 

The Composition of Coal. The uncombined carbon in coal is known as fixed carbon. Some 
of the carbon constituent is combined with hydrogen, and this, together with other gaseous sub- 
stances driven off by the application of heat, forms that portion of the coal known as the volatile 
matter. The fixed carbon and the volatile matter constitute the combustible. The oxygen 
and nitrogen contained in the volatile matter are not combustible, but custom has applied this 
term to that portion of the coal which is dry and free from ash, thus including the oxygen and 
nitrogen in the combustible. 

TABLE 1 

APPROXIMATE CHEMICAL CHANGES PROM WOOD FIBER TO ANTHRACITE COAL 



Subatanee 



Carbon 



Wood Fiber 52. «5 

Peat 69.57 

Usnite 66.04 

Earthy Brown Coal ' 78.18 

Bituminous Coal , 76.06 

Semi-Bituminous Coal ! 89.29 

Anthracite Coal I 91 .58 



Hs^QTOfen 


Ozycen 


6.26 


42.10 


5.96 


84.47 


6.27 


28.69 


6.68 


21.14 


5.84 


19.10 


5.05 


5.66 


3.96 


4.46 



Coals may be classified according to the percentages of fixed carbon and vc^tile matter 
tained in the combustible. Wm, Kent gives the following classification. 

TABLE 2 

CLASSIFICATION OF COALS 



Name of Coal 



Anthracite 

Semi-Anthracite. . 
Semi-Bituminous . 
Bituminous* East. 
West 
Lignite 



Pbrcbntagbb op Combustibub 



Fixed Carbon 



97.0 to 92.5 
92.5 to 87.5 
87.5to75.0 
75.0 to 60.0 
65.0 to 50.0 
50.0 and under 



Volatfle Matter 



S.Oto 7.5 
7.5 to 12.5 
12.5 to 25.0 
25.0 to 40.0 
85.0 to 50.0 
50.0 and over 



B.Uu. per Pound 
of Combustible 



14,600 to 
14,700 to 
16,600 to 
14,800 to 
18,600 to 
11,000 to 



14,800 
16,600 
16,000 
16,800 
14,800 
18,600 



52 



FUELS AND CX)MBUSTION 



53 



Tha non^-^ombuMie eonstUutnU are the ash and moisture, the former vai3ring from 3 per 
cent to 90 per cent and the latter from 0.75 to 25 per cent of the total weight, depoading on lo- 
cality where mined and grade. A large percentage of ash is undesirable, as it not only reduces 
the calorifie vahie of the fuel, but chokes up the air passages in the furnace and through the 
fud bed, thus preventing the rapid oombustion necessary to high efficiency. If the coal con- 
tains an exoeesive quantity of sulphur, trouble will result from its harmful action on the metal 
of the boiler where moisture is present, and because it unites with the ash to form a fusible slag 
or clinker which will choke up the great bars and form a solid mass in which large quantities 
of unccHisuined carbon may be imbedded. 

Moittwre in coal may be more detrimental than ash in reducing the temperature of a furnace, 
as it is Doo-combustible, and absorbs heat both in being evaporated and superheated to the tem- 
permture of tbe furnace gases. In some instances, however, a certain amount of moisture in a 
bituminous ooal produces a mechanical action that assists in the oombustion and makes it pos- 
sible to develop hi^ier ci^Muaties than with dry coaL 

Gtneral CharacUridicB qf Hard and Soft Coals, The (onner contain fixed or uncombined car- 
bon in large proporticm, whereas the latter have an increasing percentage of carbon in combina- 
tion with hydrogen, or hydrocarbon, which is volatile, and will distill off under high temperature, 
producing sn^oke. Hard ooal usuaDy contains more ash, especially in the smaller sizes. The 
distinguishing characteristics <A the various coals are given in the following paragraphs as de- 
scribed in " Steam," Babcock & WUcox Co. 

AnHuadte or Hard Coal. This ooal ignites slowly, but when in a state of incandescence its 
radiant heat is very great. Its flame is very short and of a yellowish blue tinge and it can be 
burned with practically no smoke. This coal does not swell when burned although it contains 
from 3 to 7.5 per cent of volatile matter. 

True or dry anthracite is characterised by few joints and clefts, and their squareness; great 
relative hardness and density; high specific gravity, ranging from 1.4 to 1.8, and a semi-metallic 
luster. 

Anthracite is now daased and marketed according to graded sizes and designations as given 
in Table 3. 



TABLE 3 

NAMES AND SIZES OF ANTHRACITE OR 



■HARD" COAL 



Namei of SixM 



Will Pan Through 



•• nSJ:::::::::::::::::::::::::::::::::::::::::::: ! >^i«-«»«* 

or Rtot Ji-in. mesh 

Pi« I H-in, mesh 

ChMtnut, or Nut I 1 H-in. mesh 

Sto^o or lUiiffo I 1 H-ia. meah 

Efc— fai thoEast 2J4-iii. mesh 

LarnEgf — Chinfo 4 -in. mesh 

SoMll Eft— Chiesfo 2H-in. mesh 

Broken, or Grata i 4 -in. mesh 



WiU Not Paaa Through 



K-hi. mesh 



l>i-in. 
Ifi-in. 
2H-in. 
2 -in. 
2H-in. 



mesh 
mMh 
mesh 
meah 
mesh 
mesh 
meah 
meah 



The anthracite coals are, with some unimportant exceptions, confined to five small fields in 
eastern Pennsylvania. 

Scmi-Anthnidte Coal. This coal kindles more readily, because of its higher content of 
vcdatile combustible, and bums mxx^ rapidly than anthracite. It has less density, hardness, 
and metallic hister than anthracite, and the average specific gravity is about 1.4. 

This coal is found in the western part of the anthracite field in a few small areas. 

Semi-Bitiiminous CoaL A softer coal than anthracite or semi-anthracite, contains more volar 
tile hydrocarbon, and will kindle more easily and bum more rapidly. It is usually free burning, 
•ad, owing to its hi^ cakwific value, very desirable iot steam-generation purposes. 

This coal is found in Pennsylvania, Maryland, Virginia, West Virginia, and Tennessee. 



54 POWER PLANTS AND REFlUGERATION 

Bitumiiioiis Coals. These coals are still softer than those described above and contain still 
more of the volatile hydrocarbons. The difference between the semi-bituminous and the bitu- 
minous coals is an important one, economically. The fcnmer have an average heating value 
per pound of combustible about 6 per cent higher than the latter, and they bum with much 
leto smoke in ordinary furnaces. The distinctive characteristic of the bituminous coals is the 
omission of yellow flame and smoke when burning. In color they range from pitch black to 
dark brown, having a resinous luster in the most compact specimens, and a silky luster in such 
specimens as show traces of vegetable fiber. The specific gravity is ordinarily about 1.3. 

Bituminous coals are either of the caking or nomHsaking variety. The former, when heated, 
fuse and swell in size; the latter bum freely, do not fuse, and are oommcHily known as/ree htBrn" 
ing coals. Caking coals are rich in volatile hydrocarbons, and are valuable in gas manufacture. 

Bituminous coals absorb moistvre from the atmosphere. The surface moisture can be re- 
moved by ordinary drying, but a portion of the water can be removed only by heating the 
ooal to a temperature of about 250^ F. 

TABLE 4 

KAMES and sizes of bituminous or "SOFT" COAL 

For "Domeatie'* soft ooala there are no uniform names and aises, but they are mariceted in the Tarioua atatea 
under ab<mt theee daaws: 

"SereMiinga" usually smaUest sizes. 

" Duff" eoes through H-inch screen. 

"No. 8 Nut" goes through 1 ^-in. serem, over f^4nch serera. 

*' No. 2 Nut" goes through 2-ineh screen, over 1 K-ineh screen. 

"No. 1 Domestic Nut" goes through 8-inch screen, over 1 H- or 2-indi screen. 

"No. 4 Washed" goes through ^-mch screen, over V^4nch screen. 

"No. 8 W«riied Chestnut" goes through 1 k^nch screen, over ^-inch screen. 

"No. 2 Washed Stove" goes through 2-incn screen, over 1 Vi-inch screen. 

"No. 1 Washed Egg" goes through 3-inch screen, over 2-inch screoi. 

"No. 8 Roller Screened Nut" goes through 1 *^inch screen, over ]4nch screen. 

"No. 2 Roller Screened Nut" goes through 2-inch screen, ovor 1 Vy-inch screen. 

"No. 1 Roller Screened Nut" goes through 3 H^ch screen, over 2-inch screen. 

*' Egg" Boes throuffh 6-ineh, over 3-Inch screen. 

"Lump or "Block" goes through 6-inch screen, or over* 

"R\m-of-Mine" in fine and large lumps. 

Pocahontas Smokeless: generally siaed as: "Nut," "Egg," "Lump," and "Mine-Run." 

Bituminous coal is far more generally distributed than any of the other coals, bmng found 
in the Appalachian field in the states of Pennsylvania, West Virginia, Maryland, Virginia, Ohio, 
Kentucky, Tennessee, and Alabama; a field nearly 900 miles in length. The eastern interior 
field includes Michigan, all of Illinois, and parts of Indiana and Kentucky. The western field 
includes Iowa, Missouri, Kansas, Oklahoma, Arkansas, and Texas. The Rocky Mountain fields 
include parts of Montana, Wyoming, Colorado, Utah, and New Mexico. The Pacific Coast fields 
are limited to small areas in California, Oregon, and Washington. 

Cannel Coal. This is a variety of bituminous coal, rich in hydrogen and hydrocarbons and 
is exceedingly valuable as a gas coal. It has a dull, resinous luster and bums with a bright flame 
without fusing. Cannel coal is seldom used for steam coal, though it is sometimes mixed with 
semi-bituminous coal where an iacreased economy at high rates of combustion is desired. The 
composition of cannel coal is approximately as follows: fixed carbon, 26 to 55 per cent; vola^ 
tile matter, 42 to 64 per cent; earthy matter, 2 to 14 per cent. Its specific gravity is approxi- 
mately 1.24. 

Names and Sizes of Cannel Coal: For fireplace, " Hand-Picked Lump "; for stoves, " Egg." 

Lignite. Organic matter in the earlier stages of its conversion into coal is known as lignite 
and includes all varieties which are intermediate between peat and coal of the older formation. 
Its specific gravity is low, being 1.2 to 1.23, and when freshly mined it may contain as high as 
50 per cent of moisture. Its appearance varies from a light brown, showing a distinctly woody 
structure, in the po(H'er varieties, to a black, with a pitchy luster resembling hard coal, in the 
best varieties. It is non-caking and bums with a bright but slightly smoky flame with moderate 
heat. It is easily broken, will not stand much handling in transportation, and if exposed to the 
weather will rapidly disintegrate, which will increase the difficulty of burning it. 



FUELS AND COMBUSTION 55 

Its composition varies over wide limits. The ash may run as low as 1 per cent and as high 
as 50 per cent. Its high content of moisture and the large quantity of air necessary for its com- 
bustion cause large stack losses. It is distinctly a low^ade fttdf and is used almost entirely in 
the districts where mined, because of its cheapness. 

Lignites resemble the brown coals of Europe and are found in the western states of Wyo- 
ming, New Mexico, Arizona, Utah, Montana, North Dakota, Nevada, California, Oregon, and 
Washington. Many of the fields given as those containing bituminous coals in the western states 
also contain true lignite. Lignite is also found in the eastern part of Texas and in Oklahoma. 

Pest. This is organic matter in the first stages of its conversion into coal and is found in 
bogs and similar places. Its moisture content when cut is extremely high, averaging 75 to 80 
per cent. It is unsuitable for fuel until dried, and even then will contain as much as 30 per cent 
mouture. Its ash content when dry varies from 3 to 12 per cent. In this country, though large 
deposits of peat have been found, it has not as yet been found practicable to utilize it for steam- 
generating purposes in competition with coal. In some European countries, however, the peat 
industry is common. 

Pressed Fuels. In this class are those fuels composed of the dust of some suitable combus- 
tible, pressed and cemented together by a substance possessing binding, and in most cases, in- 
flammable properties. Such fuels, known as hriquetteSf are extensively used in foreign countries 
and consist of carbon or soft coal, too small to be burned in the ordinary way, mixed usually with 
pitch or coal tar. Much experimenting has been done in this country in briquetting fuels, the 
government having taken an active interest in the question, but as yet this class of fuel has not 
come into common use, as the cost and difficulty of manufacture and handling have made it im- 
posBible to place it in the market at a price to compete successfully with coaL 

Coke. This is a porous product, consisting almost entirely of carbon, remaining after cer- 
tain manufacturing processes have distilled off the hydrocarbon gases of the fuel used. It is 
produced (1) from gas coal distilled in gas retorts; (2) from gas or ordinary bituminous coals 
burned in special furnaces called coke ovens; end (3) from petroleum by carrying the distillation 
of the residuum to a red heat. 

Coke is a STnokeless fuel. It readily abaorba moisture from the atmosphere and if not kept 
under cover its moistiure content may be as much as 20 per cent of its own weight. 

Qas-house coke is generally softer and more porous than oven coke, ignites more readily, 
and requires less draft for its combustion. 

• Names and Sizes of Domestic By-Product Coke: " Egg,'' 3-in. to 2H-in. " Large Stove," 
2H-in. to 2-in. "SmaU Stove,*' 2-in. to IH-in. "Nut," l^-in. to Ji-m. "Pea," ?i-in. to H-in. 

The heat values of coke range from 12,500 B.t.u. per 1 lb. to 13,500 B.t.u., depending on the 
ash content, which may vary from 5 to 10 per cent. 

Coal Analysis. The analysis of a coal should be ascertained if possible. The actual com- 
position of any coal is determined by an ultimate chemical analysis, which can only be made 
by an experienced chemist. 

The uUimale analysis of a fuel gives the percentage by weight of the various elements com- 
posing same. Such an analysis is usually reported on the dry sample as 100 per cent, and the 
percentage (^ moisture in the origihal sample given separately. 

The true analysis is easily obtained by dividing each reported percentage by 100 plus the 
percentage of H/) in the original sample as indicated in Table 5. 

The proximate analysis of a fuel gives the percentage by weight of the fixed carbon, volatile 
matter, moisture, and ash. 

The moisture is found by heating a finely pulverized sample (through a 100-mesh sieve) for 
one hour in a drying oven at a temperature of 240^ to 280^ F. The loss in weight in this time is 
due to moisture. 

The sample is then heated to a red heat for several hours in a closed crucible to expel the 
volatile matter (gases). Weighings are made at intervals and no air is allowed to come in contact 
with sample until a constant minimum weight is reached. 



POWER PLANTS AND REFRIGERATION 
TABLE 5 

TYPICAL ITLTIHATE ANALYSIS 



Cbmlat'a Repot Tnu Anilyrii 




Finally the aunple is hnt«d to a wbito beat and the fixtd earban allowed to cafobine with 
the oxygm of the air, forming carbon dioxide gaa (COi). 

llie reeidue remaining is tuHi or incombuatiiiU, and if a careful record of weighings haa been 



PlO. I. MABUn-BOUB Oalomhbtcb. 



made the lose of wei^t for each step represents successively moisture, volatile matter, Bxed car- 
bon, and the final residue, the ash. 

See Table 6 tar results (rf pr<Hdmat« analyses m Anthracite and Semi-Anthradtfis. 

Heat Valiw of a Fuel. The betd oS combutfion or oaior^c mJim q/ a /ud is the nmnba of 
B.t.u. ev(rfved when 1 lb. of the fuel is completely burned in air or oxyicen. 

A fuel calorimeUr ia used to determine the heat generated by the combustion of a kniown 
weight of the fuel, and this heat reduced to a pound basis. In the case of a aoiid or liquid fuel 
a bomb ctdorimder (Fig. 1] is employed, and the standard apparatus In use at the present time 
m essentially the some as fhat devised by M. Pierre Mahkr. 



FUELS AND COMBUSTION 



57 



In such an apparatus the fuel is completely burned, and the heat generated by the oombittr- 
tkm is abeorfoed by water, the amount of heat being calculated from the increase in the temper- 
ature of the water. A calorimeter which has been accepted as the best for such work is one in 
which the fuel is burned in a steel bomb filled with compressed oxygen. The functi<Hi of the 
oxygen, which is ordinarily under a pressure of about 25 atmospheres, is to cause the rapid and 
complete combustion of the fuel sample. The fuel is ignited by means of an electric current, 
allowance being made for the heat produced by such current, and by the burning of the fuse wire. 

The apparatus consists of: A water jacket, A, which maintains constant conditions outside of 
the ctSmixneter proper, and thus makes possible a more accur&te computation of radiation losses. 

The poroelain-lined steel bomb, B, in which the combustion oi the fuel takes place in com- 
pressed oxygen. 

The platinum pan, C, for holding the fuel. 

The calorimeter proper, D, surrounding the bomb and containing a definite weighed amount 
of water. 

An dectrode, E, connecting with the fuse wire, F, for igniting the fuel placed in the pan, C. 

A support, G, for a water agitator. 

A thermometer, /, for temperature determination of the water in the calorimeter. The ther- 
mometer is best supported by a stand independent of the calorimeter, so that it may not be moved 
by tremocB in the parts oi the calorimeter, which would render the making oi readings difficult. 
To insurs accuracy, readings should be made throu^ a telescope or eyeglass. 

A spring and screw device for revolving the agitator. 

A lever, L, by the movement of which the agitator is revolved. 

A pressure gage, A/, for noting the pressure of the oxygen admitted to the bomb. Between 
20 and 25 atmospheres are ordinarily employed. 

An oxygen tank, O. 

TABLE 6 

COMPOSITION AND HEAT VALUES OF ANTHRACITE COALS 



Locality 



Anthndie 
Ivania 

be 

\m 

m 
» 

VaDey 
VaD^ 
Valky 

ut'.V,'. 

leki 

Siemi-Antiimdte 

PtBiMUflTuiat Loyabodc 

Bwnka 

Bcrniet 

WUGMlNttTe 

Vb^^Natuna Coke 

ladiMiT^ 
Marylaiid, EasDy 



Ffzed 
Car- 
bon 


Vola- 
tile 


Mola- 
ture 


Aah 


Sul- 
phur 


78.60 




• • • • 


14.80 


0.40 


81.82 


8.84 


3.88 


10.96 


0.67 


76.94 


6.42 


1.34 


15.80 


• • • • 


79.28 


8.78 


3.88 


13.70 


• • • 




84.46 


5.87 


0.97 


9.20 


• • • 




89.19 


1.96 


8.62 


5.28 






75.20 


7.86 


1.44 


16.00 


• • • 




76.94 


6.21 


• • • • 




• • • 




81.00 


5.00 


• • » • 


• t • • • 


• • • < 




86.40 


8.08 


3.71 


6.22 


0.58 


82.66 


3.95 


3.04 


9.88 


0.46 


88.94 


2.38 


1.60 


7.11 


0.01 


87.74 


3.91 


2.12 


6.35 


0.12 


86.00 


• • • • • 


• • • • 


7.00 


0.90 


74.49 


14.73 


1.52 


9.26 


• • • • 


83.84 


8.10 


1.30 


6.23 


1.03 


82.62 


3.66 


0.96 


8.27 


0.24 


89.89 


8.56 


0.97 


9.34 


1.04 


88.90 


7.68 


t ■ • • 


3.49 


• • • • 


71.68 


13.84 


0.67 


13.96 


0.03 


76.08 


12.44 


1.12 


11.38 


0.47 


74.06 


14.98 


1.35 


9.66 


• • • • 


73.21 


13.66 


5.11 


8.03 


1.18 


88.60 


16.40 


• • • • 




• « • 





B.t.u. 

per Lb. 

of Dry 

Coal 



12,200 
11301 
12,149 
12,294 
13.728 
12,423 
15,300 
15300 
15,000 
15,070 



13317 

15,400 
15.050 
15,475 
14,199 



13,662 
11,207 



A battery or battoies, P, the current from which heats the fuse wire used to ignite the fueL 

This or a similar calorimeter may be used in the determination of the heat of combustion of 

solkl or liquid fuels. Whatever the fuel to be tested, too much importance cannot be given to 



58 POWER PIANT8 AND REFRIGERATION' 

the BMuriag of an avMage sample. Where cool ia to be tested, tests should be made from a poi^ 
tion of the dried and pulveriied laboratMy sample, the methods of obtaining which have beoi 
described. In considering the methods of calwimeter detennination, the remarks ^plied to 
coal are equally applicable to any solid fuel, and such changes in methods us are neoessary fi^ 
liquid fuels will be self-evident from tbe same description. 

A considerably simpler form of apparatus has been perfected by Profeetor S. W. Parr, which 
depends upcHi the oxidizing effect oi sodium peroxide to " bum " tbe fuel. The results are not 
Be accurate as those obtained with the Mahler apparatus, but serve for many classes of comiDncia] 

Heat values iA typical American coals are given in Tables 6 and 7 as determined by the 
Mahler-bomb calorimeter. 

TABLE 7 

HRAT VALUES OF BITUMINOUS COALS 





i»ti talnm Inim t/. £. Gfobftail S<»v BunMlB No. 8S!. ud (/. S. Ssnwi 
0/U>MBuUMinN(i.£3 


SUM 


No. 


Klndof Fud 


Coun^ 


iSc-i 


A1.h». 


BIS 

y; 

BOS 

i 
is 
i 

490 

IS 

i 

49» 

z 

BBS 

1 

290 


















3( 

1 ■■■ 










































































1 


































































































fe?^"^;:: 














5?rs^r::::.::: 










Subbit— Ine bunilni 






















gX;'".-:. :::.::. 






iSiSfSii; :;:.;: 

















NOTS.— The abova vil 
vilim of t)M varioua cmla i 



■ wen obtained at the Rt. Ltmu TaHns Pkml ti 

m EatabUabad by "actually bumJnc one iram of tl . 

■lus In B.t.u. give tbs UwaiBtleal minlmiim thermal n 






Hig^ and Low Heat Vslue of Fuels. For any fuel containing hydrogen the caknific value 
as found by the cslorimeter is higher than can be realized under roost working conditions existing 
in boiler practice by an amount equal to the latent heat of the water formed by combustitm. 
This heat would reappear if the vapor was condensed, but in ordinary practice the vapor paaes 
away uncondeoeed. This fact gives rise to a distinction in heat values between the soK»lkd 
" higher " and " lower " calorific values. The higher vahie, i.e., the one detenniDed t^ the 



FUEia AND COMBUSTION 59 

cftloriiDeta-, ia the pn^ier BCJentific unit, is the value which should be umd in boiler teating work, 
and is the one moommeDdrd by the American Socirty of Mechanical Etiginfrr*. 

TABLE 8 
HEAT VALUES OP tLUNOI3 COALS 



SaliH , 

Cra» Cnck 
VbXn 



C«tnl IlUnoii. . 

OntnUm 

BifMuddr 



12,1M 
18,197 



■ of ub about Z9 par cmU. and di 



■ bMtInc valuta about B per oant. 



Per Ceil of Find CartKA B Combuulitc 
no. 3. GKAFHicRBrBiaBNTA'noN or Relation BBTWBRN 
Heat Value Per Pound of CouBVsnaLB ako Pixro 

OABSON in COMBVnTBLE AS DEDUCED BT Wm. KbNT, 



There is no absolute measure of the tower beat 
vahie, and in view ctf the wide difTerence in opinjoa 
among physicists as to the deductions to be made 
from the hi^ier or absolute unit in this determina- 
tion, the loner value must be conaJdeted an artifi- 
cial unit. The lower value entails the use <rf an 
ultimate analysis and involves assumptions that 
would make the employment d such a unit imprac- 
ticable (or commercial work. The use oi the low 
value may also lead to error and is not recoromended 
for boiler practice. 

An example of its illogical use way be sbowa 



60 



POWER PLANTS AND REFRIGERATION 



by the consideration of a boiler operated in connection with a special eoonomiaser where the 
vapor produced by hydrogen is partially condensed by the economizer. If the bw value were 
used in computing the boiler efficiency, it is obvious that the total efficiency of the oombined 
boiler and economizer must be in error through crediting the combination with the heat imparted 
in condensing the vapor and not charging such heat to the heat value of the ooaL 

Calorific Value by Formula. The following expression known as Du Lang's fcMrmula for 
heating value per pound of coal can be used if the ultimate analysis of the fuel is known: 



F = 14,600 C + 62,000 



(«-i) 



+ 4,000 5, 



where C, H, 0, and S represent the proportionate parts of each clement per 1 lb. of fuel, and 
F denotes the heat value in B.t.u. per pound due to combustion. 

This formula does not apply when the fuel contains carbon monoxide, CO^ but can be made 
to apply by adding a term, 10,150 C, in which C is the proportionate part of carbon burned 
to monoxide. 



Example. Application of fonnula to a coal of ultimate analysis as here given follows: 

Analysis (Based on fuel as received) 



14,600 X 0.7479 + 62,000 



( 0.( 



c 


74.79% 


H 


4.98 





6.42 


N 


1.20 


8 


3.24 


H^ 


1.55 


Ash 


7.82 




100.00% 


Then by Du Long' 9 formula ' 


0.0642 \ 



8 



+ 4.000 X 0.0324 - 13,650 B.t.u. per 1 lb. ood. 



A bomb-calorimeter test showed 13,480 B.t.u. for this coal. The formula fails to allow for evap- 
orating and superheating the moisture present in the fuel. 

Heat Value Based on Fixed Carbon. The relation between the heat value per pound of 
combustible and the fixed carbon in the combustible is shown by Fig. 3 as deduced by Wm. KenL 

Calorific Value of Gaseous Fuels. The calcvlalion of the calorific value of gaseous fuds may 
be made by means of Du Lon^s formula provided the constituent gases are separated into their 
elementary gases and a term is added to provide for carbon monoxide, or the calculaticm may 
be based on the percentages of the constituent gases present and the heat value of each, as given 
in the following table: 

TABLE 9 

WEIGHT AND CALORIFIC VALUE OF VARIOUS GASES AT 82« F. AND ATMOSPHERIC PRESSURE 
WITH THEORETICAL AMOUNT OF AIR REQUIRED FOR COBiBUSTION 



Gas 



Hydroffen 

Carbon monoxide 

Methane 

Acetylene 

Olellantgas 

Ethane 



Symbol 



Cubic 
Feet of 
Gas per 

Pound 



H 

CO 

CH4 

CsHi 

C>H4 

CtH4 



177.90 
'2.81 
52.37 
13.79 
12.80 
11.94 



B.t.u. 

per 

Pound 



62000 
4460 
23550 
21465 
21440 
22230 





Cubic 


B.tu. 


Ftoetof 


cSbie 


Air Re- 
quired 


Feet 


per Pound 




of Gaa 


349 


428.25 


347 


30.60 


1053 


214.00 


1656 


164.87 


1675 


183.60 


1862 


199.88 



CabieFeet 
ofAfr 



;ubie 
Footof 
Gaa 



2.41 

2.89 

9.67 

11.98 

14.38 

16.74 



FUELS AND <X)MBUSTION 61 

Bxampto. A«8iiiiio a Datiiral gaa, the analysis of which in percentages by vohime is oxygen » 0.40, 
rarbon monoxide •- 0.95, carbon dioxide » 0.34, olcfiant gas (C2H4) « 0.66, ethane (CsHe) a 3.55. 
marsh gaa (CH4) "■ 72.15. and hydrogen » 21.95. All but the oxygen and the carbon dioxide are 
oombustiblee, and the heat value per cubic foot will be: 

From CX) « 0.0095 X 347 - 3.22 

C1H4 " 0.0066 X 1675 - 11.05 

CsH« » 0.0355 X 1862 - 65.99 

CH4 « 0.7215 X 1053 - 757.58 

H - 0.2195 X 349 - 75.95 

B.t.u. per cu. ft. == 913.79 

The net air required for combustion of one cubic foot of the gas will be: 

CX) « 0.0095 X 2.39 - 0.02 

CtHi - 0.0066 X 14.33 - 0.09 

CtH« - 0.0355 X 16.72 - 0.59 

CH4 - 0.7215 X 9.54 - 6.88 

H - 0.2195 X 2.39 - 0.52 

Total net air per cu. ft. - 8.10 

LIQUID FUELS— OIL 

PetroIomiL Hie following distinguishing characteristics of petroleum have been taken 
from " Steam," Babcock A Wilcox Cor. 

"Petroleum is practically the only liquid fuel sufficiently abundant and cheap to be used for 
the generatioii of steam. It possesses many advantages over coal and is extensively used in 
many localities. 

" There aire three kinds of petroleum in use, namely, those yielding on distillation t 1st, paraf- 
fin; 2Dd, asphalt; 3rd, olefine. To the first group belong the oils of the Appalachian Range and 
the Middle West of Uie United States. These are a dark brown in color with a greenish tinge. 
Upon their distillation such a variety of valuable light oils are obtained that their use as fuel is 
prohibttiye because of price. 

" To the second group belong the oils found in Texas and California. These vary in color 
from a reddish brown to a jet black and are used very largely as fuel. 

** The third group comprises the oils from Russia, which, like the second, are used largely f<Mr 
f uel purpoaee. 

" The light and easily ignited constituents of petroleum, such as naphtha, gasoline, and kero- 
mie, are oftentimes driven off by a partial distillation, these products being of greater value 
for other purposes than for use as fuel. This partial distillation does not decrease the value of 
petroleum as a fuel; in fact, the residuum known in trade as fud oil has a slightly higher 
calorific value tiian petroleum and because of its higher flash point, it may be moro safely 
handled. Statements made with reference to petroleum apply as well to fuel oiL 

" In general, crude oil consists of carbon and hydrogen, though it also contains varying quan- 
^tiee of mmsture, sulphur, nitrogai, arsenic, phoeph(vus, and silt. The moisture contained may 
vary from less tiian 1 to over 30 per cent, depending upon the care taken to separate the water 
from the oil in pumping from the well. As in any fuel, this moisture affects the available heat 
of the oil, and in contracting for the piutshase of fuel of this nature it is well to limit the percentage 
of mcHsture it may contain. A large portion of any ccmtained moisture can be separated by . 
settling and for this reason sufficient stc^age capacity should be supplied to provide ~time for 
such action." 

Tlie ealerifie valves qf pdrokum range from 18,000 to 22,000 B.t.u. per pound, and the per- 
eeotage composition and other data are given in Table 10. The flash point of crude oil ia-the 
temperature at which it begins to give off inflammable gases. This temperature varies greatly 
for different oil8» as shown in the table. 



62 



POWER PLANTS AND REFRIGERATION 

liberated in sufficient quantity 



Tbe fin point m the t«mperatuT« fit which these gaaes 
to bum continuously. 

TABLE 10 " 

COMPOSITION AND CALORIFIC VALUB OP VARIOUS OIU 



Kind ot CHI 


Fa 

cfti. 




p« 

Cent 
phur 


Ovtn 


C^mSty 


B 


B.t.u. 


ADthdrity 




if 

ii'.s 

iJ;? 


ill 

M 

ifii 


il 

i'.r 

1.8 


OM- 


M 


870 


iVMl 

E 

19680 

iiiio 






6 

j 




























B.4W.C0. 


















ffiiJ."-;.;:: 















■ K. 



tP< 



tUquMPMlBoMd. 



The comporoltDe talue tff pelrolewn and coal a» fuel may be suouned up to the advantage of 
the liquid fuel bb follows: The cost of handling ia much lower, both in delivery and in buniing 
same, while for equal heat value much less storage space is required, and this qwoe may be at 
a distance ftom tike Ixnlera. Higher efficiencies are obtainable, since the combustion is mcfe 




Via. i. PXABODT Oil Bcrncb. 

perfect, lees excess air ia required, temperatures are more constant, and ainoe smtAe is larg^ 
eliminated, the heating surfaces axe correspondingly clean. 

The int«neity ct tbe Gre can be instantly regulated to suit the load requirements, and there 
is no deterioration from less of heat value by disintegration due to storage. 

The duadanUage of Ihe liquid fuel aiises from the fact that the oil must have a reasofiably 
high flash point to reduce the danger <^ explosion, and city ordinances may, in certain cases make 
its use practicaUy prohibitive. Owing to the high temperatures of the oil flame tbe boiler up- 
keep cost may be increased. 

The comparative evsfiorative power of ooal and oil is given in Table 11 



FUELS AND COMBUSTION 



63 



TABLE 11 

EVAPORATION OF WATER FROM COAL AND OIL 
TUbbb from the •< U. 8. Geolosieia Report on Petroleum " for 1900 



Dedgnatlon of Coal 



Non.— One ton eoel 
One barrel oO 
OnegaDonoO 



-2.0001b. 

- 42 fala. or 886 lb. 

-81b. 



P itteburg lump and nut, Pemwylvania 
Phtaburi nut and alaek. Pennaylvania 

Anthracite, Pennegrlvanlm 

fiMJiMM BkM^ 

O eo rat e Ctmk. lump, Maryland 

Neir River, West Vfrcinin 

Pocahontas lump. West Virfinia \ 

Cardiff hmp, Waloa 

Cape Breton, Canada 

Naaafano, Britiah Columbia 

Britiah Columbia 
Greta. W) 

wm, WaahinctoQ 



Pounds of 


Barrels of 


Water 


Petroleum 


Evaporated 


Required 


fitMn and 


to do same 


at212<'per 


Amount of 


Pound of 




Combus- 


tionasl 


tible in 


Ton of Coal 


the Coal 


Petroleum 




18*' to 40«» 




Baum6 


10.0 


4.0 


8.0 


8.2 


9.8 


8.9 


9.6 


8.8 


10.0 


4.0 


9.7 


8.8 


10.6 


4.2 


10.0 


4.0 


9.2 


8.7 


7.8 


2.9 


8.9 


8.6 


7.6 


8.0 


7.6 


8.0 



Under favorable conditions 1 pound of oO will evaporate from 14 to 16 pounds of water from and at 212®; 1 pound 
of coal wfll evaporate from 7 to 10 pounds of water fitnn and at 212®; 1 pound of natural gas will evaporate from 18 
to 20 pounds of water from and at 212®. 



on Bomiiig. The burning of petroleum fuel or oil can only be accomplished in steam-boiler 
practice by the use of suitable burners, which must atomize the oil so thoroughly that each par- 
tide will be brought into contact with the minimum quantity of air necessary for its complete 
cmnbustion before the gases come in contact with any heating surfaces. The furnace must be 
of hi^ily refractwy material, the radiant heat from which will assist in the combustion. No local- 
isatioQ of the heat must occur at the heating surfaces or trouble will result {torn overheating 
and Mistering. 

The hwmen may be classified under three general t3rpos: 1st, spray bumen, in which the oil 
IB atomised by steam or compressed air; 2nd, tfapor bttmers, in which the oil is converted into 
vapor and then passed into the furnace; 3rd, mechanical burners, in which the oil is atomiied 
by submitting it to high pressure and passing it through a small orifice. 

The Peabody Bttmer (Fig. 4) is of the latter type. These mechanical burners have been in 
general use only a short time in this country, and the round-flame burner has proved more satis- 
factory than the flat-flame burner of this type. 

The efficiency of oil burning with boilers of 500 horsepower may run as high as 83 per cent 
grosB or 81 per cent net after deducting 2 per cent for steam used by burner. The conditions of 
average practaoe are such that efficiencies ranging from 5 to 10 per cent less than the above are 
about the best that may be expected. 

GASEOUS FUELS 

The gaseous fuels in most oommcm use are blast furnace gas, natural gas, and by-product 
ooke-oven gas. 

Blast Pomace Gas. This is a by-product from the blast furnace of the iron industry; the 
oomposition of a typical sample from a Bessemer Furnace is as follows: 

OOt - 10.0%, 00 = 26.2, H - 3.1, CH4 = 0.2, N - 60.6. 

With the exoepticMi of Uie small amount of carbon in combination with hydrogen as methane, 
tod a yery small percentage of free hydrogen, ordinarily lees than 0.1 per cent, the calorific value 



G4 



POWER PLANTS AND REFRIGERATION 



of blast furnace gas is due to the CX) content which when united with sufficient oxygpa as used 
under a boiler, finally bums to COs. The heat value of such gas will vary in most cases frnn 
85 to 100 B.t.u. per cubic foot under standard conditions. In modem practice, where the blast 
is heated by hot blast stoves, approximately 15 per cent of the total amount of gas is used for 
this purpose, leaving 85 per cent of the total for use under the boilers or in gas engines, thai is, 
approximately 8500 pounds of gas per ton of pig iron produced. In a modem blast fumaoe 
plant, the gas serves ordinarily as the only fuel required. 

Natural Gas. This gas has a limited use but is, <^ course, confined to restricted areas. 
The best results are secured by using a large number of small burners to which the gas is supplied 
at a pressure of about 8 ounces. The calculations for amount of gas required to give a certain 
heating effect should in all cases be based on volume reduced to standard ccmditions of tempera- 
ture and pressure, namely, 32^F., and 14.7 lb. pressure per sq. in. 

The variation in composition and heating value of natural gas is shown in the following table: 

TABLE 12 

TYPICAL ANALYSIS (BY VOLUME) AND CALORIFIC VALUES OF NATURAL GAS 

FROM VARIOUS LOCALITIES 



Locality of Well 



Andenon, Ind . 

Findlay, O 

St.Ive, Pa 

Pittsburgh, Pa. 
Pittaburgh, Pa. 















Heavy- 




H 


CH« 


CO 


COi 


N 





HydnH 
Carbona 


BS 


1.86 


W.07 


0.78 


0.26 


8.02 


0.42 


0.47 


0.15 


1.64 


98.86 


0.41 


0.25 


3.41 


0.89 


0.86 


0.20 


6.10 


75.54 


Trace 


0.34 


• • • • • 


• « • • 


18.12 


• • • • 


9.64 


57.86 


1.00 


• « • • 


28.41 


2.10 


6.00 




20.02 


72.18 


1.00 


0.80 




1.10 


4.30 


« • • • 



B.tai. 
per Cu. 
FtlCal- 
eulated* 



1017 
1011 
1117 

748 
917 



* B.t.u. calculated, using peromitagea of constituent gases, and separate heat values. 

By-pfoduct Coke-Oven Gas. This is also known as artificial ffos, or iUuminating gas, and 
is a product <^ the destructive distillation of coal in a distilling or by-product coke oven. In 
this class of apparatus the gases, instead of being burned at the point of their origin, as in a bee- 
hive or retort coke oven, are taken from the oven through an uptake pipe, cooled, and 3rield 
as by-products: tar, ammonia, and illuminating and fuel gas. A certain pc^rtion of the gas 
product is burned in the ovens and the remainder used or sold for illuminating or fuel purposes, 
the methods <^ utilizing the gas varying with plant operation and locality. 

Table 13 gives the anal3r8es and heat value of certain samples of by-product coke-oven gas 
utilized for fuel purposes. 

This gas is nearer to natural gas in its heat value than is blast furnace gas, and, in general, 
the remarks as to the proper methods of burning natural gas and the features to be followed in 
furnace design hold as well for by-product coke-oven gas. 



TABLE 13 

TYPICAL ANALYSIS OF BY-PRODUCT COKE-OVEN GAS 





Sample No. 


COt 





CO 


CH« 


H 


N 


B.t.u._per 
Cu.Ft. 


1 


0.76 
2.00 
8.20 
0.80 


Trace 

Trace 

0.4 

1.6 


6.0 
3.2 
6.3 
4.9 


28.16 
18.80 
29.60 
28.40 


53.0 
57.2 
41.6 
54.2 


12.1 
18.0 
16.1 
10.1 


505 


2 


899 


8 


551 


4 


460 







The essential difference in burning the two fuels is the pressure under which it reaches the 
gas burner. Wliere this is ordinarily from 4 to 8 ounces in the case of natural gas, it is approii- 



FUELS AND COMBUSTION 



05 



inately 4 inches of water in the case of by-product coke-oven gas. This necessitates the use of 
larger gas openings in the burners for the latter class of fuel than for the former. 

By-product coke-oven gas comes to the burners saturated with moisture, and provision should 
be made fcH* the blowing out of water of condensation. 

CX)MBUSTION 

Comlrastion of Fuel. Ck>mbustion as used in steam engineering signifies a rapid chemical 
combination between oxygen, and the carbon, hydrogen and sulphur composing the various 
fuds. This combination takes place usually at high temperature with the evolution of light 
and heat. 

The substance combining with the oxygen is known as the combustible, and if it is completely 
burned or oxidised the combustion is perfedt that is, no more oxygen can be taken up by the 
products of the reaction. 

The combustion is imperfecl or incomplete when carbon bums to form carbon monoxide, CO, 
instead of the dioxide, CX)i, since the former may be further burned to form carbon dioxide if the 
necessary oxygen is supplied. 

The temperature at which the reaction begins to take place is known as the kindling tenu' 
perature and is different for each combustible. The following values are from Stromeyer: 





TABT.K 14 

KINDLING TEMPERATURES 




Fuel 


Temp.F. 


* 

Fuel 


Temp. P. 


linite dust 


800«»P. 

486 

470 

670 

600 


Ck>ke 


Red Heet 


Dried peat 


Anthracite 


Red He«t 760<'F. 


Sulphur 




Red He«t 1211 


Anthracite duat 


Hydrogen 


1080-1290 


Coel 











CkMnbustion takes place only between hot gases and oxygen, hence all ccnnbustibles are prac- 
tically gaseous at the instant of combustion. 

The characteristics of these gases and atmospheric air must be definitely known before 
combustion problems can be solved, and such data will be found in the following tables: 

TABLE 15 

DENSITY OP GASES AT 32« P. AND ATMOSPHERIC PRESSURE 29.92 INS. 

(ADAPTED PROM SMITHSONIAN TABLES) 



Gm 


Chemical 
Symbol 


Specific 
Gravity* 
Air -1 


Weight 

of One 

Cubic 

Foot, 

Pounde 


Volume 
of One 

Pound. 

Cubic 

F^et 


RBL4TIVB 

DBNsnv, 
Hydbogbn b 1 




Exact 


Appr. 


Oxygen 


O 
N 
H 
COi 
CO 
CHi 
CsH« 
CsHa 
SOt 


1.063 

0.9678 

0.0696 

1.6291 

0.9672 

0.6676 

1.076 

0.920 

2.2689 

1.0000 


0.08922 
.07829 
.006621 
.12269 
.07807 
.04470 
.08879 
.07264 
.17862 
.08071 


11.208 
12.778 
177.90 

8.161 
12.809 
22.871 
11.936 
13.786 

6.698 
12.890 


16.87 
18.92 

1.00 
21.88 
18.89 

7.96 
14.91 
12.91 
81.96 


16 


Nhrosen 


14 


Hydrogen ^ ^ 


1 


f^firbon dioxide . 


22 




14 


MuthflT^ 


8 


F'thene 


15 


Acetylene 


13 


Sulphur dioxide 


32 


Air ::::...;.;:: 











Combustion Reactions. The constituent elements of a gas combine with oxygen in {K^r- 
fectly definite proportions by weight and volume, forming definite combusHon producU, These 
reactions as well as the proportions in which the gases combine have been tabulated for use in 
computation work and are given herewith. 



66 



POWER PLANTS AND REFRIGERATION 



TABLE 16 

OXYGEN AND AIR REQUIRED FOR COMBUSTION 

Bt Wbigrt 

At 32® F. and 29 . 92 Inches 



1 


2 


8 


4 6 


6 


7 


8 


9 10 

1 


Oxidi>- 
aUe Sub- 
stance 
or Com- 
bustible 


Chemi- 

eal 
Symbol 


Atomk 

or 
Com- 
bining 
Wgt. 


Cbemiral 
Reaction 


Product of 
Combustion 


Oxygen 

per 

Pound 

ofCoU 

umnl in 

Pounds 


Nitrogen 
per 
Pound 
of Col- 
umn 1 " 
8.82*XO 

in 
Pounds 


Air 

per 

Pound 

ofCoU 

umnl- 

4.82tXO 

in 
Pounds 


Gaseous 
PftMhaet 

Pound 
of Col. 
1« 1 + 
Col. 8 
in Lbs. 


Beat 

Vahie 

Pouil 

ofCoL 

1 in 

B.t.a. 


Carbon. . . . 
Carbon.... 
Carbon 
monoxide 
Hydrogen.. 
Methane... 

Sulphur. . . 


C 
C 

CO 

H 
CH4 

S 


12 
12 

28 

1 
16 

82 


C-l-20-COt 
C+O -CO 

CO+O-CO, 
2H+0-HsO 
CH4+40- 

COt+2HiO 
S+20-80s 


Carbon dioxide... 
Carbon monoxide. 

Carbon dioxide. . . 

Water 

Carbon dioxide 

and water 

Sulphur dioxide . . 


2.667 
1.888 

0.671 
8.0 

4.0 

l.o 


8.85 
4.48 

1.90 
26.66 

18.28 
8.82 


11.62 
6.76 

2.47 
84.66 

17.28 
4.82 


12.62 
6.76 

8.47 
86.66 

18.28 
6.82 


14600 
4460 

lOlfiOt 
62000 

28660 
4060 



* Ratio by iveight of N to O in air. 

J 4.82 pounds « air contain one pound of Q. 
Per pound of C in the CO. 



TABLE 16 (CanHnued) 
Bt Volumb 



1 


2 


11 


12 


18 


14 


15 


16 


17 


18 


Oxidixable 


Chemical 


Vol- 
umes 
of Col- 
umn 1 
Enter- 
Com- 
bination 
Volume 


Vol- 
umes of 
Oxygen 

Com- 
bining 

with 
Column 

11 
Volume 


Volumes 
Product 
Formed 
Volume 


Volume 

per Lb. 

01 Column 

1 in Gas- 

eotts Fomif 

Cu.Ft. 


Volume 
of Oxygen 
per 
Pound 
ofCoU 
umn 1, 
Cu.Ft. 


Volume 
of Pro- 
ducts of 
Combus- 
tion per 
Pound 
of Col- 
umn 1» 
Co. Ft. 


Volume 
of Nitro- 

Pound 

ofCoU 

umn 1 — 

8,782»X 

Column 

16kCtt.Ft. 


Vohmc 

of Gasnr 

Pound 


Substance or 
Combustible 


Symbol 


1 -Col- 
umn 16 + 

17«.Ca.Ft. 


Carbon 

Carbon 

Carbon monoxide 

Hydrogen 

Methane 

Sulphur 


C 

C 

CO 

H 

CH4 

S 


I 
1 
2 
2 

1 
1 


2 
1 
1 
1 
2 
2 


2COt 

2CO 

2COt 

1C(>S^^ 
2SOt 


14.95 
14.96 
12.80 
179.82 
22.41 
6.60 


29.89 
14.96 
6.40 
89.66 
44.88 
11.21 


29.89 
29.89 
12.80 
179.82 
67.84 
11.21 


112.98 

66.49 

24.20 

889.09 

169.65 

42.89 


142.87 
86.88 

87.00 
618.41 
286.89 

68.60 



* Ratio by volume of N to O in air. 



B^beoek A Wtkoac Co. 



It will be seen from this table that a pound of carbcm will unite with 2^ pounds of oxygen 
to form carbon dioxide, and will evolve 14,600 B.t.u. As an intermediate step, a pound of 
carbon may unite with IH pounds of oxygen to form carbon monoxide and evolve 4460 B.t.u., 
but in its further conversion to COi it would unite with an additional IH times its weight of 
oxygen and evolve the remaining 10,150 B.t.u. 

When a pound of CX) bums to CX)i, however, only 4350 B.t.u. are evolved, since the pound 
of 00 contains but */i lb. carbon. 

Air Required for Combustion. It has already been shown that each combustible element 
in the fiiel will unite with a definite amount of oxygen. With the ultimate anal3rBi8 of the fuel 
known, the theoretical amount of air required for combustion may be readily calculated. 



FUELS AND COMBUSTION 67 

Bxample. Let the ultimate analysis be as follows: per Cent. 

C Tbon. 74.79 

nyovofQiu 4.08 

Oxytcn 6.42 

Nitroftti 1 . 20 

Mphur. 8.24 

Water - 1.66 

Aflli 7.82 

100.00 
When complete combustion takes place, as already pointed out, the carbon in the fuel unitce with 
a definite amount of oxygen to form COt. The hydrogen, cither in a free or combined state, will unii4* 
with oxygen to form water vapor, HtO. Not all of the hydrogen shown in a fuel analysis, however, in 
available for the production of heat, as a portion of it is already united with the oxygen shown by the 
analyaia in the form of water, HiO. Since the atomic weights and H and O are respectively 1 and 16, 
the weight of the combined hydrogen will be )^ of the weight of the oxygen, and the hydrogen available 
for combustion will be H ~ V$ O. In complete combustion ol the sulphur, sulphur dioxide, SOt,i8 formed. 
Expressed numerically, the theoretical amount of air required for the above analjrsis is as follows: 
(See Column 6, Table 15.) 

0.7479 C X 2% « 1.9944 O 

(0.0498 - 0.0642/8) H X 8 - 0.3262 O 
0.0324 S X 1 » 0.0324 O 

Total required 2.3610 O 

One pound of oxygen is contained in 4.32 lb. of air. 

The total air needed per pound of coal, therefore, wUl be 2.3610 X 4.32 « 10.200 lb. 

The wei^t of combustible per pound of fuel is 0.7479 + 0.0418* + 0.0324 + 0.012 « 0.83 pounds, 
and the air theoretically required per pound of combustible is 10.200/0.83 » 12.3 lb. 

The above is equivalent to computing the theoretical amount of air required per pound of fuel 
by the formuUt: (See Ck>lumn 8, Table 16.) 

Weight per pound « 11.53 C + 34.56 (H - 0/8) + 4.32 S - 
where, C, H. O and S are proportional parts by weight of carbon, hydrogen, oxygen, and sulphur by 
ultimate analjrsis. 

Theoretical and Actual Amount of Air Required. The calculations for air required presup- 
pose that each and every particle of oxygen can be brought into intimate contact with the com- 
bustible. Practically this is impossible, due to the large amount of inert nitrogen present, varia- 
tioDs in the fuel bed, and interference of clinker and ash, which cannot be removed as soon as 
formed. When burning oil and gas, however, some of these difficulties are eliminated, and the 
actual can more nearly approach the theoretical amount of air as calculated and given in Table 17. 

TABLE 17 

THEORETICAL AMOUNT OP AIR REQUIRED 



V9I. V 


COMPOSmON BT WnOBT 


Lbs. of 
Ai^per 


rud 


%c 


%H 


%o 


Lb. of 

FtMl 


Wood diareoal 


98.0 
80.0 
84.0 
91.5 
87.0 
70,0 
68.0 
60.0 
86.0 


■ • • • 

• ■ • • 

• • • • 

3.6 
6.0 
6.0 
6.0 
6.0 
18.0 


• • • • 

• • • • 

• • • • 

2.6 

4.0 

20.0 

81.0 

48.6 

1.0 


11.16 




9.6 


Cikt 


10.8 


A— IffgfifM Qoal ....... 


11.7 


Bfl^lHMantfnfl aqaI^ AFV.. ... .... 


11.6 


ij??l *^^ 


8.9. 


BtB^^- oFV .. . • . . ......•••.•....■• 


7.68 


wKdTSri ... . 


6.00 


M^i?l 


14.80 







It 18 therefore necessary to provide for an excess of air when burning coal under either nat- 
ural or forced draft, amounting to approximately 50 to 100 per cent of the net calculated amount, 
or about 18 to 24 lb. per pound of cdal. 

Less air results in imperfect combustion and smoke, while an excess cools the fire and setting 
and carries away large quantities of heat in the flue gases. 
♦Avaaafale hydrogn. 



68 POWER PLANTS AND KEFRIGERATION 

FLUE GAS ANALYSIS 

Composition of Flue Gas. A flue gaa analysis gives the propcn-tion by volume of the priji- 
cipal constituent gases produced by the combustion of any fuel. The gases usually determined 
in such an analysis are carbon dioxide, CO3, oxygen, O, and carbon monoxide, CO, whilp the 
residue or volume remaining after these gases are removed is taken as nitrogen, N. 

By reference to Table 16 it will be seen that when oxygen and carbon combine the volume of 
the carbon dioxide gas formed is exactly equal to the volimiie of oxygen entering into the reac- 
tion, provided all volumes are measured at the same temperature and pressure. It, therefore, 
follows that if just sufficient air is provided to bum exactly one pound of pure carbon, the gas 
resulting will contain 20.91 per cent CX)) and 79.09 per cent N, the oxygen having all entered 
into combination with the carbon, and the new gas resulting has simply taken the place of the 
original 20.91 per cent O. Now if 50 per cent excess air is supplied only % of the original oxy- 
gen volume will be replaced by COj and the flue gas analysis will show 13.91 per cent CX>j, 7.0 per 
cento and 79.09 per cent N. Finally, if 100 per cent excess air is supplied only J^ <rf the original 
oxygen volume will be replaced by CO, and the flue gas will contain 10.45 per cent OOi, 10.45 
per cent O, and 79.09 N. In each case the oxygen or sum of the oxygen and carbon dioxide per^ 
centage is constant or 20.91 per cent, while the nitrogen percentage is likewise constant at 79.09 per 
cent provided pure carbon only is burned completely. 

If carbon monoxide is produced it will occupy twice the volume of the oxygen entering into 
its composition, hence the volume of the flue gas resulting will be greater (at the same tempera- 
ture and pressure) than that of the air supplied by J^ of the per cent of CO present. One volume 
C + one volume O = two volumes CO. 

If hydrogen is present in the fuel it will increase the apparent percentage of nitrogen in the 
flue gas, due to the fact that the water vapor formed by its combustion will condense at the tem- 
perature of the analysis, while the nitrogen brought in with the oxygen which combined with 
the hydrogen will remain as a gas and appear in the analysis. 

Actual Air Supplied for Combustion. Likewise the total or actual amount of air supplied per 
pound of fuel burned can be expressed as follows, provided the flue gas analysis is known, and 
the relative densities of the gases are given. 

These densities are in the same ratio as the molecular weights, which are as follows: COs » 
44, CO = 28, O, = 32, N, = 28; in which C = 12, O = 16 and N = 14. 

In this connection it must be remembered that equal polumes of all gases at the same tem- 
perature and pressure contain the same number of moleculeSf hence the truth of the above statement. 

It will thereCore be apparent that if we let Ni, COi and CO represent the percentages by 
volume from a flue gas analysis, and Ci the percentage by weight of carbon in the fuel; then 
the pounds of air per pound of fuel will be expressed as follows: 

A 28 X N, 

• 12(C0, + C0) X76.9 ' 

where 76.9 » per cent of nitrogen in atmospheric air by weight and A^ « lb. of air supplied per 
pound of the fueL 

It should be noted that in the above expression all the carbon is supposed to bum and pass 
up the flue. Since this is never true in practice, it is necessary to correct Ci by the amount of 
carbon in the ash. Thus, if the ash in a boiler test amounted to 16 per cent, and an anal3rsi8 was 
found to contain 25 per cent of carbon, the percentage of unconsumed carbon would be 16 X 0.25 «= 
4 per cent of the total coal burned. Now if the coal by ultimate analysis contained 80 per cent 
of carbon, only 80 — 4 = 76 per cent of the fuel woukl actually be combustible carbon, hence 
use 76 per cent for Ci in the above formula instead of 80 per cent, which is Ci, as repwted in 
the analysb. 

Then the ratio of air actually supplied to that theoretically required is A^/At as determined above. 



FUEIB AND COMBUSTION 69 

Waight «i Fine Qu. The terigU qf Jlue dum, W, per pound of eailxm i* mko easlj oNDpatei 
ffom tbe flue gaa analysis by the following formuU, 

44 CO, + 32 O. + 28(CO + N) 
12 (CO, + CO) ' 

wheta the aymbcds COi, O,, CO and N are the perceotagee by volume of these gases as detei^ 
mined from the flue gas aoalysiB. Abo the weight of flue gas per pound of dry oool may be de- 



'IjMS due to Meal carried cway ty Cbimnsy Gases for VOrying 
Percenioges of Corbon OlCKlcle 
Fio. 0. 

tertnioed from tliis formula by multiplying IF by the percentage of carbon Ci in tbe coal as found 
by an ultimate aoalysiB. 

Heat Loat in Flue Gu. Ttm heal toil in Ike five {)ama due to the heat in the gates ia L = 0.24 
W {It — li) *diere L — B.t.u. lort per pound rf dry coal, W = weight of flue gases per pound 
of dry coal, (, — tenip««ture of flue gases, !■ = temperature of air, and 0.24 = specific heat of 
Uw Sue gases. The above loss is given graphicaUy, as shown by Fig. 6, for varying percentages 
of OOt and different flue gas temperatures. 

The heat lottin the fiue gases, due to the formaUon of carbon monoxide when tiie carbon is iif 
completely burned is, in B.t.u. per pound of dry fuel, 

^■■"■■""< 12(roTcO.) '"^" 

where 10,150 is the heat value per pound of carbon in the CO, and CO and CO, are percentages 
by vohime from the flue gas analysis while Ci is the proportion by weight of carbon which must 
be corrected to give the unount bumod and passed up the stack as already explained. 

Omt ^par«tns. The apparatus most commonly used for flue gas analysis is known as the 
Onat (Fig. 7), and is described as follows: 

" Tlie burette A is graduated in cubic centimeters up to 100, and is surrounded by a wat«r 
jacket to prevent any change in temperature from affecting tbe density of the gas being analysed. 

" fteeecunUewOTk it is advisable to use four pipettes, fi,C, D, £, the fint containing a sohi- 



70 POWER PLANTS AND REFRIGERATION 

tioD of eauMe jxitatk /or the abtorption of earion dioxide, tbe aeoond an alkiiline w^tioa of pyn- 
gaUol for Oie abtorption <^ o^vgm, and tbe remRining two an acid ■alution of euprout ehiaride for 
dbtorbing ihe carbon moruaide. Each pipette oontains a. nmnber of glass tubeo, to which some 
of tbe solution clings, thus racilitatiug the abaocpticHi of the gas. In tite pipettes D and B, cop- 
per wire ia placed in these tubes to re-energiie tbe 
solution BB it becomes weakened. The rear half of 
each pipette is fitted with a rubber bag, one of which 
is shown at K, to protect tbe solution from tbe ac- 
tion of the ur. The solution in each pipette should 
be drawn up to the mark on the capillary tube, 

" The gaa is drawn into the burette through tbe 
IJ-tul>e B, which is filled with spun glass, or similar 
motaial, to clean tbe gas. To discbarge any air or 
gaa in the apparatus, the cock G is opened to the air 
and the bottle P is raised until the water in the bu- 
rette leacbes tbe 100 cubic-centimeter msxk. Tbe 
cock G ie then turned ao as to close the air opening 
and allow gas to be drawn through H, the bottle F 
being lowered ftn* this purpose. The gaa is drawn 
into the burette to a pmnt below tbe sero mark, the 
cock then being opened to the air and the excess 
gas expelled until tbe level of the wat^ in F and in 
A is at tbe zero mark. This operation is neceesary 
in order to obtain the tero readily; at atmoqiheric 



"The apparatus should be carefully (ufed/or^foA-- 
age aa wdl as all conncctiona leading thereto. Simple 
tests can be made, as for example: If after the cock ^__^ 

Q is closed, the bottle F is placed on top (A the frame Pio. 7. Obsat li 

for a abort time and again brought to the sero marie, 
and the level of the water in il ia above tiie aerO maib, a leak is indicated. 

" Before taking a final sample for analysis, the burette A should be filled with gaa and emp- 
tied (Hioe or twice, to make sure that all the apparatus is filled with the new gas. The cock G 
is then closed and the cock I in the pipette B is tqwned and the gas driven over into B by raising 
the bottle F. The gas is drawn back into A by lowering F and when tbe solution in B haa reached 
tbe mark in the capillary tube, the cock / is closed and a reading ia taken on the burette, the 
level erf the water in the bottle F being brought to the same level aa the water in A. The opera- 
tion is repeated until a constant reading is obtained, the number of cubic centimeten, absorbed 
aa shown by the reading, being the percentage of COi in the flue gases. 

"The gaa ia then driven over into the pipette C and a similar operation is carried out. Hk 
difference between the resulting reading and the first reading gives the percentage of oxygen io 
the flue gaaes. 

" The next operation is to drive the gas into the pipette D, tbe gas being given a final wash in 
£, and tben paased into the pipette C to neutraliae any hydrochlcHic acid fumeg which may have 
been given <rft by the cuixx>us chloride solution, which, eepeciaUy if it be oM, may give off sudi 
fumes, thus increasing the volume of the gners and making the reading on the burette kaa than 
tbe true amount. 

" The proeeaa must be carried out in the order named, as the pyrogallol solution will aln 
absorb carbon dioxide, while the cuprous chloride solution will also absori) oxygen. 

" Aa the pressure of the gases in the flue is lees than the atmospheric pressure, they will BOt 
<d themeelv<i) flow through the pipe connecting the Buc tu the apparatus. The p« may bs 
drawn into the pipe in the way already described for filling the ^pantusi but this is a ledioai 



FUELS AND COMBUSTION 71 

method. For rai»d work a robber bulb aspirato connected to the air outlet of the ooek win 
enable a new supply of gas to be drawn into the pipe, the apparatus then being filled as already 
described. Another form of aq>iratQr draws the gas from the flue in a constant stream, thus in- 
suring a firesh supply for each sample. 

" The analysis made by the Orsat apparatus is volumetric. If the ancdysia by weight is re- 
quired, it can be found from the volumetric analysis as follows: 

" Multiply the percentages by volume by either the densities or the molecular weight of each 
gas, and divide the products by the sum of all the products; the quotients will be the percen- 
tages by weight. For most w<»^ sufficient accuracy is secured by using the even values of the 
molecular wei^ts." 

Kzample. An application of the above data when an ultimate analysis of the fuel and a volumetric 
aoalysiB of the flue gas is known can be made as follows: 

Partial ultimate analysis, C - 82.1%. H - 4.25%. O - 2.6%, S - 1.6%. Ash - 6.0%. and 
B.t.u. per pound of dry Pocahontas coal « 14,500. The flue gas analysis is, 

Percent 

COt 10.7 

9.0 

CO 0.0 

N (by difference) 80.3 

• 

Determine: The flue gas analysis being given, (1) the amount of air required for perfect combus- 
tion. (2) the actual weight of air per pound of fuel, (3) the weight of flue gas per pound of coal, (4) the 
heat lost in the chimney gases if the temperature of these is 500® F., and (5) the ratio of the air supplied 
to that iheoretieally required. 

Sobition: The theoretical weight of air required for perfect combustion, per pound of fuel, from 
formula already given under ** Air Required for Combustion.*' will be. 

(0.026\ 
0.0425 — 1 + 4.32 X 0.016 - 10.88 lb. 

If the amount of carbon which is burned and passes away as flue gas is 80 p^ cent, which would allow 
for 2.1 per cent of unbumed carbon in terms of the total weight of dry fuel burned, the weight of dry 
gas per pound of carbon burned will be from formula already given imder "Weight of Flue Gas." 

^ 44 X 10.7 + 32 X 9.0 + 28 (0 + 80.3) ^, ^^ ,. 

Ft 1 ■• _ ■» Zo.4ifi ID.| 

12 (10.7 + 0) 

and the weight of flue gas per pound of coal burned will be 0.80 X 23.42 « 18.74 lb. 

The heat lost in the flue gases p^ pound of coal burned will be from formula and the value 18.74 
iost determined: 

Loss - 0.24 X 18.74 X (500-60) » 1.979 B.t.u. 

The parentage of heat lost in the flue gases will be 1,979 X 100/14.500 » 13.6 per cent 

The ratio of air supplied per pound of coal to that theoretically required will be (18.74 — 1)/10.88 « 
1.63. 

The ratio oi air supplied per pound of combustible to that required will be. 

0-803 

— — ^— ^— — ^-^— — — ^^^^— — — Bi 1 73 

0.803 - 3.782 (.09 + >4 X 0) 

N % nitrogen in whole amoimt of air. 

* N - 3.782 (O + J^ CO) " % nitrogen in air actually required. 

NoTB. The value 3.782 is the volumetric ratio of nitrogen to oxygen in the air (Table 16. Colunm 
17). An the unoombined oxygen and H of the carbon monoxide represents the oxygen equivalent of 
nnneefwwi y or excess nitrogen, which in turn represents air. 

The ratio based on combustible will be greater than the ratio based on fuel if there is unconsumed 
carbon in the ash. 



73 POWER PLANTS AND REFRIGERATION 

Unreliability of CDs Readings Taken Alone. It is generally assumed that high COt read- 
iogs are indicative of good combustion and hence of high efficiency. This is true only in the 
sense that such high readings do indicate the small amount of excess air that usually acoom- 
panics good combustion, and for this reason high COi readings alone are not considered entirely 
reliable. Wherever an automatic COi recorder is used, it should be checked from time to time 
and the analysis carried further with a view to ascertaining whether there is CO present. As 
the percentage of COt in these gases increases, there is a tendency toward the presence of CO, 
which, of course, cannot be shown by a COt recorder, and which is often difficult to detect with an 
Orsat apparatus. The greatest care should be taken in preparing the cuprous chloride solution 
in making analyses and it must be known to be fresh and capable of absorbing CO. 

Smokeless Combustion. Smokeless combustion can only be attained with special equip- 
ment and most careful firing, and the following methods for its accomplishment are reconmiended 
by the Babcock A Wilcox Co., who have had a wide experience in this field: 

'' The question of smoke and smokelessness in burning fuels has recently become a very im- 
portant factor in the problem of combustion. Cities and communities throughout the country 
have passed ordinances relative to the quantities of smoke that may be emitted from a stack, 
and the failure of operators to live up to the requirements of such ordinances, resulting as it do^ 
in fines and annoyance, has brought their attention forcibly to the matter. 

'^ The whole question of smoke and smokelessness is to a large extent a comparative one. 
There are any number of plants burning a wide variety of fuels in ordinary hand-fired furnaces, 
in extension furnaces and on automatic stokers that are operating imder service conditions, prac- 
tically without smoke. It is safe to say, however, that no plant will operate smokeleady under 
all condUions of service^ nor is there a plant in which the degree of smokelessness does not depend 
largely upon the intelligence of the operating force. 

'' When a condition arises in a boiler room requiring the fires to be brought up quickly, the 
operaiives in handling certain types of stokers wiU use their slice bars freely to break up the 
green portion of the fire over the bed of partially bmned coal. In fact, when a load is suddenly 
thrown on a station the steam pressure can often be maintained only in this way, and such use 
of the slice bar will cause smoke with the very best type of stoker. In a certain plant using a 
highly volatile coal and operating boilers equipped with ordinary hand-fired furnaces, extension 
hand-fired furnaces and stokers, in which the boilers with the different tjrpes of furnaces were 
on separate stacks, a difference in smoke from the different types of furnaces was apparent at 
Ught loads, but when a heavy load was thrown on the plant, all three stacks would smoke to 
the same extent, and it was impossible to judge which type of furnace was on one oar the- other 
of the stacks. ^ 

"In hand-fired fiamacea much can be accomplished by proper firing. A combination ot the 
alternate and spreading methods should be used, the coal being fired evenly, quickly, lightly, 
and often, and the fires worked as Uttle as possible. Smoke can be diminished by giving the 
gases a long travel imder the action of heated brickwork before they strike the boiler heating 
surfaces. Air introduced over the fires and the use of heated arches, for mingling the air with 
the gases distilled from the coal will also diminish smoke. Extension furnaces will undoubtedly 
lessen smoke where hand-firing is used, due to the increase in length of gas travel, and the fact 
that this travel is partially under heated brickwork. Wliere hand-fired grates are inunediately 
under the boiler tubes, and a highly volatile coal is used, if sufficient combustion space is not 
provided, the volatile gases, which are distilled as soon as the coal is thrown on the fire, strike 
the tube surfaces and are cooled below the burning point before they are wholly consumed and 
therefore pass through as smoke. With an extension furnace, these volatile gases are acted 
upon by the radiant heat from the extension furnace arch, and this heat, together with the added 
length of travel, causes their more complete combustion before striking the heating surfaces than 
in the former case. f 

" Smoke may be diminished by employing a baffle arrangement which gives the gases a fairly 
long travel under heated brickwork and by introducing air above the fire. In many cases, how- 



FUELS AND COMBUSTION 73 

ever, special furnaces for smoke reduction are installed at the expense of capacity and 
eamomy. 

'* Fnun the standpoint of smokelessness, undoubtedly the best results are obtained with a good 
stoker, properly opo^ted. As stated above, the best stoker will cause smoke under certain oondir 
tions. Intelligently handled, however, under ordinary operating conditions, stoker-£red fur- 
naces are much more nearly smokeless than those which are hand-fired, and are, to all intents 
and purposes, smokeless. In practically all stoker installations there enters the element of time 
for oombustion, the volatile gases as they are distilled being acted upon by ignition or other 
arches before they strike the heating surfaces. In many instances, too, stokers are installed 
with an extension beyond the boiler front, which gives an added length of travel, during which, 
the gases are acted upoa by the radiant heat from the ignition or supplementary arches, and 
here again we see the long travel giving time for the volatile gases to be properly consumed. 

" Finally, it must be emphatically borne in mind that the question of smokelessness is largely 
one of degree, and dependent to an extent much greater than is ordinarily appreciated upon the 
handling of the fuel and the furnaces by the operators, be these fmnaces hand-fired or automati- 
cally fixed." 



CHAPTER IV 

BOILERS AND RULES FOR CONSTRUCTION 

POWER BOILERS 

The term " power boiler," as generally understood, refers to a boiler in which a pressure of 
approximately 80 lb. per sq. inch or more is employed for supplying steam to various types of 
prime movers. The term " heating boiler " refers to boilers designed to carry only a low pressure, 
usually not over 25 lb. per sq. inch working pressure, for supplying steam to low pressure heat- 
ing or drying systems. A boiler installation involves, among other thingiB, a consideration of 
the following items: calorific value of the fuel, grate area, draft, and boiler heating surface. 

Grate Area. In order to generate a definite weight of steam at a certain pressure (or evapoiv 
ate a definite weight of water) in a unit of time (lb. per hour) from water at a given temperature 
requires a fixed amount of heat to be supplied by the combustion of some fuel. That is, a definite 
number of pounds of fuel must be burned per hour, depending upon the calorific value of the 
fuel| to supply the heat required. The amount of fuel that is burned on a square foot of grate 
surface per hour is termed the rate of cambusium and is limited by the character of the fuel, 
draft, etc. Therefore, it may be said, with more or less exactness, that for a certain weight of 
water to be evaporated per hour under certain conditions of pressure, feed-water temperature 
and calorific value of the fuel to be used, a definite amount of grate area will be required. 

Draft To bum a given weight of fuel of certain character in a unit of time requires a definite 
amount of air to supply the oxygen necessary to support combustion, as was previously shown to 
be the case under " Fuels and Combustion." The air is passed under the grate and through the 
fuel-bed and meets with more or less resistance both in passing through the fuel-bed and later 
through or around the boiler tubes and flue. It is necessary, then, to supply a motive force to 
circulate the air required either by a fan, steam jet or chimney. The absolute pressiu^ existing 
over the fuel-bed must be less than the absolute pressure existing under the grate to cause a 
Aoiw. The difference between the atmospheric pressure and the pressure existing at any point 
through the fiunaoe or the flue connecting the boiler with the chinmey or st^uik is termed the 
"draft" at that particular point. This pressure difference is measured and stated in inches of 
water. The question of draft is fully treated in the Chapters on ** Chimneys for Power Boilere " 
and " Mechanical Draft." 

Bdler Heating Surface. The amount of heat that may be transmitted from the hot gases 
through a unit area of the steel shell or tubes of a boiler to the water in a unit of time is limited 
by the temperature difference between the gases and water, and the velocity of the gases over the 
heating surface. 

This difference is limited by the temperature as obtained from the combustion of the fuel 
and the temperature of the water in contact with the surface. Consequently, a boiler, in order 
that it may absorb the necessary heat to evaporate a given weight of water in a unit of time, 
must be supplied with a definite amount of surface, termed ** heating surface," in contact with 
the hot gases. 

From the foregoing statements it is evident that the character and calorific value of the 
fuel, grate area and draft and boiler heating surface are dependent upon one another. The 
consideration of any boiler installation will, therefore, involve the calculation or assumption of 
the magnitude of the items mentioned from the data obtained in a laboratory, and from actual 
tests of existing plants. 

Heat Transfer in Boiler Tubes* There have recently been completed, under the direction 

74 



BOILERS AND RULES FOR CONSTRUCTION 



75 



ci J. E, BeUf an extensive and comprehensive set of experiments on ** Heat Transfer Rates."* 
The apparatus used consisted of a 2-in. internal diameter copper pipe, surrounded by 20 individual 
water jackets each 1 ft. long. The gases w^e drawn from an illuminating-gas furnace in which 
temperatures above 2,600 deg. F. could be obtained. The range of ga»-flow rate covered was 
from 4,000 to 14,000 lb. per hr. per sq. ft. of crossHsectional area of passage. The difference in tem- 
perature between the gas and metal surface was from 400 to 2,000 deg. F., and the temperature 



20 



18 



one 
I 

f 

« 

03 
I 
S 9 








■™^^ 


























































— 


Correct to within 2% for §2 In, Intornal 
diamoter tube with • wall tamp, of 180° F, 

The atraight llnaa, howavar, ara probably 
tanganta to ourvaa whhh, aa tha weight of 
gaa Inoraaaaat bend downward. 












V. 












yy 


















rfl 










* 


>7 




%t 


























yy 


























#y 


/y 




^y 
























b*V 


yy 




^y 






















J 


vy 


yy 




9 






















1A% 


yy- 


/y 


yrj/^ 






















7A 


yy 


/, 




P 




















/j 


y/ 


V/ 


'/ 


Jfa 




















/A 


// 




^y 


-J^ 


V' 




















//./ 


yy 




yj 


'^^ 


















/^^ 


y/ 


yU 




/^ 


















, 




VA 


y/ 


y/A 


Y 




















/a 


y/. 


yy 


A' 


^"^ 






















<^4*. 


yy. 


yy 


y^ 


r* 






















<^6 


V. 


'^y^ 


w^ 
























YA 


X' 


Af^ 






















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vvv 


y.'' 


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wy, 


yy 


























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. 








^ 


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— u 


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^ 


























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y 


























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^^ilH 


























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V 


1 






















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~R»fH 


e-Co\ 


arad- 


fn-fn 


DtfW/y 


ante- 








;^ 


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' 






















r^— 























































































































































OIM. 2000 4000 6000 8000 tOOOO 12000 
Weight of Qaaea per Sq. Ftof Rue Area par Hour 

Fig. 1. Heat Tranbfbb in Boiler Tubes* 



14000 LU 



o£ the metal surface varied from 145 to 215 deg. F., the average wall temperature being 180 deg. 
The variation in specific heat of the gases was taken into account. 

The flue gases, after leaving the experimental tube, were cooled in a 2-in. coil 26 ft. long, 
surrounded by wat^. The gases leaving the coil were passed through a box, where the dew 
point was determined. The dew point, together with the entering and exit temperatures of the 
gas and water through the cooler and the flue-gas anal3rses taken during the tests, gave the most 
accurate method of determining gas weights. 

The rate of heat transfer may be expressed by the formula, 



R = A+B 



(v) 



^The Bobeoek A Wilcox Co., Bayonne, N. J. 



76 K)WER PLANTS AND REFRIGERATION 

where 

R » the rate of heat transfer; B.t.u. per sq. ft. per hour per degree difiPerenoe in temperature, 

A = a constant, 

B ^ & function of the temperature difference, 

W 

— = the rate of mass flow per unit area of channel. 

a 

The value of A, as determined by these experiments, is 2.20. The value of B varies with 
changes in the temperature difference from 0.000770 for a temperature difference of 400 deg. F. 
to 0.001120 for a temperature difference of 2,000 deg. F. These values of B may be readily 
obtained from the chart for any rate of gas flow and any temperature difference. 

Fig. 1 gives in graphic form the results of these experiments. 

It is interesting to note that while at high rates of gas flow the temperature difference has 
an important bearing on the heat-transfer rate, at low weights of gas flow, such as are encountered 
in boiler practice, the effect of the temperature difference is relatively small. 

Relation Between Gas Temperaturei Heating Surface Passed Over and Amount of Steam 
Generated. Fig. 2, reproduced from ''Steam,'' shows the relation of gas temperatures, heating 
surface passed over and work done by such surface for use in cases where the temperatures ap- 
proach those found in direct-fired practice, and where the volume of gas available is approxi- 
mately that with which one horsepower may be developed on 10 square feet of heating surface. 
The curve assumes what may be considered standard gas passage areas and, further, that there 
is no heat absorbed by direct radiation from the fire. 

Experiments have shown that this ciuve is very nearly correct for the conditions assumed. 
Such being the case, its application in waste heat work is clear. Decreasing or increasing the 
velocity of the gases over the heating surfaces from what might be considered normal direct- 
fired practice — that is, decreasing or increasing the frictional loss through the boiler — will in- 
eredse or decrease the amount of heating surface necessary to develop one boiler horsepower. 
The application of Fig. 2 to such use may best be seen by a consideration of the following case. 

Assume the entering gas temperatures to be 1470 degrees and that the gases are cooled to 
570 deg. F. From the curve, under what are assumed to be standard conditions, the gases 
have passed over 19 per cent of the heating surface by the time they have been cooledi470 degrees. 
When cooled to 570 degrees, 78 per cent of the heating siu-face has been passed over. The work 
done in relation to the standard of the ciuve is represented by (1470 — 670) -;- (2500 — 600) — 
45 per cent. (These figures may also be read from the curve in terms of the per cent of the work 
done by different parts of the heating surfaces.) That is, 78 per cent — 19 per cent = 59 per 
cent of the standard heating surface has done 45 per cent of the standard amount of work. 69 -f45 
» 1.31, which is the ratio of surface of the assumed case to the standard case of the curve. Ex- 
pressed differently, there will be required 13.1 square feet of heating surface in the assumed case 
to develop a horsepower as against 10 square feet in the standard case. 

Boiler Horsepower (b.hp.) The commercial rating of a power boiler is stated in terms 
of boiler horsepower. A boiler horsepower (A, S. M. B. standard) is equal to an evaporation 
of 34.5 pounds of water from and at 212® F. per hour; that is, the boiler having the feed water 
coming to it at 212° F. must furnish the necessary heat to turn 34.5 pounds of water at this 
temperature and atmospheric pressure into steam at the same temperature and pressure per 
hour. In other words: The boiler must supply the latent heat of vaporization or 971.7* B.t.u. 
to each pound of water to turn it into steam at atmospheric pressure (14.7 lb. per sq. in. absohite) 
or it must furnish 971.7 X 34.6 = 33,523.7 B.t.u. per hour to the water per boiler horsepower. 

Therefore, 1 boiler horsepower = 33,523.7 B.t.u. per hour. 

The horsepower developed by a boiler in operation is determined by first finding the B.t.u. 



* Marks and Davis steam tables give 9704 for the latent beat, oorresponding to 212*^ F., in wbieb ease 1 bJip. 
88478.8 B.t.u. per bour. 



BOILERS AND RULES FOR CONSTRUCTION 



I 

n 

a 

S t 

^ 
1 

I s 

li 
H 

M 

I j 



78 POWER PLANTS AND REFRIGERATION 

» 

received by the water and steam from the boiler per hour and dividing this quantity by the 
B.t.u. equivalent of one boiler horsepower. 

Steam may exist in the three following states: 

1st. Saturated steam with suspended moisture. 

2d. Dry saturated steam. 

3d. Superheated steam. 

Steam with x parts vapor means that each pound of the vapor will carry {\-x) parts of water 
held in suspension. This water has not received the latent heat necessary to turn it into vapor, 
and must therefore be figured accordingly. 

Let q = heat in the water above 32^ F. corresponding to the temperature and pressure at 
which it was turned into steam. 
% s the fractional part of the nuxture of vapor and water that is vapor, 
r a latent heat of vaporization corresponding to the temperature and pressure. 
q\ = heat of the liquid, above 32** F., of the feed water. 
W s total weight of feed water furnished the boiler per hour. 
i s temperature of satturated steam corresponding to the pressure. 

Total heat, above 32^, per lb. of steam leaving the boiler: 
= g 4- X r for wet steam. 
= g + r =» H f or dry saturated steam. 
= H + C^w, {if — for superheated steam. 

The boUer horsepower developed by the boiler will be: 

b.hp. = ..^ ,^» ., — i^ steam is wet. 

^ 33,523.7 

"■ ^^ ^ _ — if steam is dry and saturated. 
33,523.7 

= ^ * if steam is superheated. 

33,523.7 

if » actual temperature of the steam if superheated. 
Cp^ B mean specific heat of superheated steam for the given range of temperature and 
pressure. See diagram, Chapter on " Water, Steam, and Air." 

Equivalent Evaporation. It is customary to refer or reduce the actual evaporation of a 
boiler to a standard set of conditions in order to make comparisons. This standard is the 
amount of water that would have been evaporated into dry steam from and at 212® F. for the 
same heat expenditure. 

W = actual wei^t of feed water per hour per lb. of fueL 
W\ = equivalent evaporation per hour. 
971.7 = latent heat corresponding to a temperature of 212^ 

W\ a» — for wet steam. 

971.7 

■= — ^» _ for dry steam. 

971.7 

= — — for superheated steam. 

Factor of Evaporation — (F). The " factor of evaporation " is the ratio of the heat required 
to generate one pound of st^am for the Riven condition (at the boiler pressure, temperature and 



X 



• 



r . 



BOILERS AND RULES FOR CONSTRUCTION 79 

feed-water temperature) to the amount of heat required to generate one pound of dry steam 
from and at 212'' F. 

Then the factor of evaporation is: y 

" — for we steam. 



97L7 



^ for dry steam. 
971.7 ^^ 

H + Cp^ (Is - 

"= tStz for superheated steam. 

971.7 *^ 

The equivalent evaporation from and at 212^ is: 

Wi^ FXW, 

Boiler Efficiency (^). The tom ''efficiency/' when applird to steam boiler performance, 
ordinarily refers to the over-all efficiency of the grate, furnace and boiler when solid fuels are 
used. 

It is obvibusly unfair to charge against the boiler the unconsumed fuel that drops through 
the grate alid becomes mixed with the ashes. It is difficult, however, to separate the efficiency 
of the boiler and furnace from the grate efficiency, and as the user must pay for any such loss 
it is customary, unless otherwise noted, to state the combined efficiency rather than separate 
efficiencies. 

It is recommended that in asking for guarantees of boiler performance when the term ''effi- 
ciency" is used that it be clearly defined in the proposal. 

When liquid fueU are used, 

^..,«,. ,. ., • , ./x Heat absorbed per pound of fuel. 

Combmed efficioicy of boiler, furnace and grate (^) » 



Calorific value of one pound of fuel. 

Wken solid fuda are used, 

^ . . , ^ . ^ . ., - , ^ V Heat absorbed per pound of fuel as fired. 

Combined efficiency of boiler, furnace and grate (^) = -^r-. — — ; ;; TTT — ; — ;: — r 

Calorific value of one pound of fuel as fired. 

Tlie efficiency of the boiler alone is stated as: 

Heat absorbed by the boiler per pound of combustible burned on the grate. 
Calorific value of one pound of combustible as fired. - 

Let C = calorific value of the fuel as fired per lb. 
B = weight of fuel per hour, lb. 
W\ ^ FW equivalent evaporation, lb. per hour. 

Then the heat absorbed per lb. of fuel consumed per hour is: 

F XW X 971.7 
B 

^ FXW X 971.7 „ F XW X 971.7 

Then ^ « or B = . 

BXC ^XC 

The equivalent evaporation per hour (from and at 212^) per lb. of fuel is: 

Wi FW 4>XC 



w = 



B B 971.7 



80 



POWER PLANTS AND REFRIGERATION 



The combined efficiency (^) that maj' be expected from the combination of well-designed 
and proportioned boilers, furnaces and grates is given in the table of tests accompanying. The 
results were practically all obtained under test conditions, and the nearness to which actual 
operations may approach these results will depend largely upon the intelligent supervision given 
to the plant. 

In the general run of plants the all-year round combined efficiency does not exceed 60 per 
cent. This is the figure usually used in preliminary estimates for small and modium-si«ed plants. 
There is practically no difference in the efficiencies of the various types of first class boilers on the 
market when inteUigently handled. 

TABLE 1 

BOILER TESTS 
B. A W. Water Tube Boaera 



Rated 

Capacitor 

B.hp. 



119 
156 
218 
300 
160 
821 
640 
300 
608 
608 
300 
298 



Coal 
Used 


Calorific 
Value 

of Coal 
B.t.u. 

per Lb. 


Method 

of 
Firing 


Ratio 
H. S. 

to 
Grate 
Area 


Draft, 
In. Water 


Per 

Cent 

Capacity 


Com- 
bined 
Effi- 
ciency 

Cmt 


In 
Furnace 


At 
Damper 


A.E. 
A.P. 
A.B. 
A.B. 
B.L. 
^G.C. 
S. 

H.O. 
M. 
C. 

P.N.S. 
A.S. 


13.464 
12,861 
11.104 
11,913 
12.292 
14,966 
14.381 
12.436 
10.676 
13.126 
13.610 
12.060 


H.P. 
H.P. 
M.S. 
H.P. 
H.P. 
H.P. 
H.P. 
M.S. 
M.S. 
M.S. 
M.S. 
H.P. 


46 

39 

42.2 

36.7 

66.6 

61.7 

64.2 

• • • • 

49 
66.4 
66.6 
60 


0.33 
.33 
.65 
.41 
.10 
.25 
.44 
.22 
.62 
.68 

1.64 
.36 


• • « « 

0.43 
• 6G 
.21 
.24 
.35 
.68 
.36 
1.24 
1.16 
.64 
.59 


84.4 
104.7 
126.7 
118.7 
101.8 

99.3 
129.3 
130.7 
161.6 
215.7 
112 
107.2 


69 

69.2 

72.1 

71.8 

71.6 

72.7 

73.2 

73.4 

74.9 

71.9 

74.6 

69.6 



Grate 
Surface 
Square 

Feet 



26.6 
40 
61.6 
84 
27 
52 
118 

• • • • « 

103.5 
90 
53 
59.6 



* The heating surface in the above boilera is equal to the rated capacity X 10. 

H.F. — ^Hand fired. S. — Somerset, Pa., bituminous. 

M.S. — Mechanical stoker. H.O. — Hocking Valley, O.. bituminous. 

A.E. — ^Anthracite egg. M. — Maacouth, 111., bituminous. 

A.P. — ^Anthracite pea. C. — Cartersville, III., bituminous. 

A.B. — ^Anthracite ouckwheat. P.N.S. — ^Pittsbw^ nut and dadc bituminoua. 

B.L. — ^Bituminous lump, Ohio. A.S. — ^Arkansas slack bituminous. 

G.C. — Georges Creek bituminous. 



Relation Between Efficiency and Capacity. When a boiler is forced beyond its normally 
rated capacity the efficiency is ordinarily somewhat decreased, although at not a very marked 
rate up to 50 per cent overloads. This is due primarily to the fact that in order to obtain a higher 
rate of evaporation the combustion rate must be increased, which in turn generates a large volume 
of flue gases. A point is soon reached where the heating surface is insufficient to absorb the 
extra heat generated, the gases leaving the boiler at a higher temperature resulting in a lowering 
of the efficiency. 

The curve Fig. 3 — A was plotted from the results of a large number of tests run on water 
tube boilers. The general direction of the curve will be found to hold approximately correct for 
operating conditions when used as a guide to what may be expected. 

Example. A certain coal gives up by perfect combustion, 14,500 B\t.u. per pound (calorific value). 
Assume a combined efficiency of 60 per cent for the boiler, furnace and grate. 

Then if all the heat from the coal was transferred to the water, the equivalent theoretical evapora* 
tion from and at 212^ F. would be: 



14,500 
971.7 



— 14.9 lb. water per lb, of coal. 



\ 



BOILERS AND RULES FOR CONSTRUCTION 81 

But since oiUy 60 per cent of the heat in the coal is transferred to the water, the equivalent evaporu- 
tion will be: 

14.500 X 0.60 

■—-— ■» 8.96 lb. water per lb. of coal. 

971.7 

Suppoae the conditions were not standard, but the steam pressure was 100-1 b. gago and the feed- 
water temperature 60® F., the heat required to raise the temperature of 1 lb. water from 60° F. and 
to convert into dry steam at 100-lb. pressure will be; 

H -qi ^ 1190.7 - 28.1 - 1162.6 B.t.u. 

14 500 X 0.60 8700 

Then — - — ,^ ^ — « « 7.6 lb. water actually evaporated per lb. of coal burned for the 

1162.6 1162.6 ^ »^ ^ 

assumed conditions of pressure and temperature of feed water. 

Fig. 3 will be found convenient for rapidly solving problems relative to boiler performance. 

Assumed Feed-Water Temperature for Estimates. It is customary to assume a feed-water 
temperature of approximately 60° F. when no feed-water heater is to be used, which is an unusual 
Gondition; and a temperature of 200° to 210° F. when a feed-water heater is to be installed in a 
non-condensing plant and approximately 175° F. for a condensing plant. This temperature will de- 
pend upon the amount of steam used by the auxiliaries. (See Chapter on " Feed Water Heaters.") 

Heating Surface — (H.S.). The heating surface of a boiler is that part of the boiler which 
has water in contact with the surface on the one side and hot gases on the other side. Super- 
heating surface is that part of the boiler having steam on the one side and hot gases on the other. 

BitUdera* Rating. It was customary for the majority of stationary boiler manufacturorn 
in the past to base the conmiercial horsepower rating of their product on the following allowance 
of heat surface per b.hp. 

TABLE 2 



Water Tube lyps 10 8.45 

Return Tubular TVpe • 12 2.88 

Seoteh Marine Type 8 4.80 




• Equiv. Evapn. 
Per Sq. Ft. H. S. 



This is known as " builders' rating." It is now the customary practice to rate all types of 
boilers on a basis of 10 sq. ft. per b.hp. 

Engineers are rapidly ceasing to rate boilers in horsepower since there is no definite re- 
lation between the rating so expressed and the horsepower of the engine which it is capable of 
driving. Moreover, there is a tendency to force boilers to evi4x>rate more water per square foot 
of heating surface. For these reasons it is preferable to speak of boHers in terms of their healing 
surface and not in horsepower rating. In selecting boiler equipment, designing engineers usually 
deteimine the rate of evaporation which they can expect per unit heating surface with the fuel, 
draft Mid setting which is to be employed. The combined water rate of the steam consumers 
is then computed and divided by the evaporative rate, which has been chosen to obtain tlio 
total heating surface of the boilers which will be required. 

Equivalent Evaporation per Square Foot of Heating Surface. The equivalent evaporation 
(from and at 212°) per sq. ft. of heating surface per hour (Wt) is equal to: The total equivalent 
evaporation per hour (W\) divided by the total amount of heating surface, or: 



H.S. 



Thus, if a boiler is rated by the builder on a basis of 10 sq. ft. of heating surface per b.hp., the 
equivalent evaporation required (from and at 212°) per sq. ft. per hour {W^) is 34.5/10 or 3.45 lb. 



82 



POWER PLANTS AND REFRIGERATION 



A boiler rated as indicated in the preceding table should develop at least 33H P^ cent more 
than its rated capacity when hand fired, using a fair grade of coal and with a draft of not kss 
than H" water available at the boiler damper, llie " builders' rating *' is simply a statement 
made by the manuf actiirer that his product under ordinary operating conditions will easily derdop, 
and with good economy, one boiler horsepower for the amount of heating surface as g^ven. It 
does not indicate the limit of actual evaporation or boiler horsepower that may be devdopecL 
This method of rating was adopted, primarily, for reasons of convenience in selling. 

When there is sufficient draft to bum the necessary fuel, water tube boilers will readily de- 
velop 200 per cent rating with a good grade of fuel. 

D, S, JacobiM states that whether the plant be hot water or steam apparently makes little 
difference as to the general oonditicms affecting the boiler capacity required. 

Apparently, with eastern coal or coal with low ash content it is possible to operate boilos 
at 200 per cent overload for a considerable period of time. With coals of a quality as is mined 
in Illinois and Indiana, 200 per cent of load may be considered as an extreme capacity. With 
western fuels, such as Ugnite, 150 per cent of load is probably the extreme which should be con- 
sidered. With oil-burning furnaces and installations, 300 per cent of load should be considered 
as the extreme capacity. 



TABLE 3 

BOILER HEATING SURFACE AND HORSEPOWER FOR RETURN TUBULAR BOILERS 



Diameter 

of Boiler, 

Inches 



Tubes 



Hbating Surface 



Length, 
Feet 



54 


14 


64 


16 


64 


14 


64 


16 


64 


14 


64 


16 


60 


16 


60 


18 


60 


16 


60 


18 


60 


16 


60 


18 


66 


16 


66 


18 


66 


16 


66 


18 


66 


16 


66 


18 


72 


16 


72 


18 


72 


20 


72 


16 


72 


18 


72 ' 


20 


72 


16 


72 


18 


72 


20 


78 


16 


78 


18 


78 


20 


78 


16 


78 


18 


78 


20 


78 


16 


78 


18 


78 


20 


84 1 


18 


84 


20 


84 


18 


84 


20 


84 


18 


" 1 


20 



DiametM*, 
Inches 



3 
3 
3H 

4 
4 
8 

k 

4 

4 
3 
3 
8H 

3H 

4 

4 

3 

3 

3 

ZH 

4 
4 
4 

3 
3 
3 

3Ki 

3H 
4 

4 

4 

3 

3 

SH 

3H 

4 

4 



Number 



Tubes 



64 
64 
44 
44 

34 

34 

72 

72 

60 

60 

46 

46 

94 

94 

70 

70 

66 

66 

118 

118 

118 

94 

94 

94 

70 

70 

70 

140 

140 

140 

108 

108 

108 

88 

88 

88 

172 

172 

136 

136 

106 

106 



662 

681 

626 

601 

467 

634 

841 

946 

684 

770 

722 

812 

1,098 

1,286 

967 

1,077 

878 

988 

1,878 

1.660 

1,722 

1,286 

1,446 

1,606 

1,098 

1,286 

1,372 

1.636 

1,839 

2,043 

1,477 

1.662 

1.846 

1,380 

1.663 

1.726 

2,260 

2,611 

2.092 

2.324 

1,871 

2.078 



SheU 



99 
113 

99 
118 

99 
118 
125 
141 
126 
141 
126 
141 
188 
166 
188 
166 
188 
166 
161 
170 
189 
161 
170 
189 
151 
170 
189 
163 
184 
204 
163 
184 
204 
163 
184 
204 
198 
220 
198 
220 
198 
220 



Rear 
Head 



8 
8 
8 
8 
8 
8 
10 
10 
10 
10 
9 
9 
11 
11 
11 
11 
11 
11 
13 
18 
13 
18 
18 
18 
18 
18 
13 
15 
16 
15 
16 
16 
15 
14 
14 
14 
17 
17 
17 
17 
16 
16 



Total 



669 

762 

688 

722 

674 

665 

976 

1,097 

819 

921 

866 

962 

1.247 

1.402 

1.106 

1.244 

1.027 

1,166 

1.642 

1.783 

1,924 

1.449 

1.629 

1.808 

1,262 

1,418 

1,674 

1,818 

2.038 

2,262 

1.666 

1.861 

2.065 

1.667 

1.761 

1.948 

2.476 

2.748 

2.807 

2,661 

2,085 

2.814' 



66 

75 

68 

72 

67 

65 

98 

110 

82 

92 

86 

96 

126 

140 

110 

124 

108 

115 

164 

178 

192 

145 

168 

181 

126 

142 

167 

181 

204 

226 

165 

186 

206 

156 

175 

194 

247 

275 

281 

266 

206 

281 



BOILERS AND RULES FOR CONSTRUCTION 



83 



In deBigning Uie boiler installation for 200 per cent <rf load, the extreme capacity, it will 
probably be fouild good practice, if not the best practice, to design a boiler to operate most 
economically at 125 to 150 p» cent of load. A condition of this kind requires larger c(»nbustion 
diambers than would be the practice if the boilers were designed to operate most efficiently at 
afullkMuL It would also require somewhat larger gas passages and also a larger stoker equiiHnent. 



.Obtoininff lUpM 
fandPwfeet 
Combuatktt with 
High 
Twuperai 




Fumaoe 




GratM 



Hich 



Air Supply 





No ooding surlaoe should be adjacent to 
or cover the fud.maae or gaaee aridog 
from fuel. 

The fuel maae and gaoes ahould be sur- 
rounded and covered by highly heated 
and radiating surteoe. 

Adequate capacity to ^ve time for per- 
fect admixture of gaa and air and 
•ecure temperature uuLiuiiiiiry for 
ehonical union. 



.Proper amount and aiae for kind of fuel* 
dnft intenaity» and boiler heating 
surface. . ^ . 

Proper air space to admit auffieient air 
yet prevent fine coal sifting through to 
ashpit. 

Proper form to permit of speedily and 
thoroughly deUvering coal to grate, 
stoking fire with fire dooca open the 
least possible time. 

Proper form to permit of speedily and 
tTOroughly removing ash with the 
least power expended or time of fire 
doors open. 



The passage of air into and through the 
fuel mass should be perfectly under 
control by damper or otherwise. 

Proper amount of air admitted to insure 
thorough combustion of gases, result- 
ing in mgh COi, low O, and no CO. 

Properly admitted to secure complete 
admixture with sases, resulting in no 
hydrogmi or hymticarbon loss, hence 
no smoke. 



Beet Flow of 
Gae to Assist 



of He 




Cause products <tf combustion to flow from above downwards 
to the greatest possible extMit. 

Cause gases to move with good velocity over and ^^ngly 
scrub heating surfaces mm. entrance to release without 
slMMrt-drcuiting. 
-Do not permit gases to stagnate at any point from entrance 
upon heating surface to rdeaae from it. 

Keep gases well broken up and sweeping the entire heating 
surface completely, not partially. • . . ^ 

Do not doaely concentrate gases on some parts of the heating 
surface and cut by others unbathed, or short-drcuit. 

From the highest furnace temperature to a rdeaae tempera- 
ture slightly above steam pressure there should be a uni- 
form drop in temperature of products of combustion. 



Beat Arrangement 
of Heating Sur- 
face to Promote 
Abstraction and 
Durability 




Use of good water free from foaming ten- 

Rapid, smooth, pre-determined water dr- 

culation. 
Circulating areas sufficient for easy down- 

wardnow of wateri -— — 

Steam space large enough to prevent etnux 

causing drop in pressure. ■ — 



Steam outlet considerable distance from ^ 
ter line and over quietest part of water. 



DurabiUty 



Requlflites which Control tfie Operation of Boilers at High Rates of Evaporation. The 
aboye chart was prepared by E, C. Fishery Pres., Wickes Boiler Co., giving in a condensed 
form the requisites which, in boiler plants, secure and maintain high rates of evaporation. 



84 POWER PLANTS AND REFRIGERATION 

Heating Surface of Return Tubular Boilers. In order to compare hcnijKHital tubular boikn 
directly both as to heating surface and horsepower, the Hartford Steam Boiler Inepection 
Insurance Company has prepared the foregoing table. It is figured on the basis of 10 sq. ft. 
of heating surface per boiler horsepower, the heating surface as calculated including the inside 
tube area, one-half the area of the cylindrical portion of the shell, and two-thirds of the area of 
the rear head minus the combined cross-sectional area of the tube. 

Grate Surface- (G) and Rate of Combustion- (/?). To evaporate a given amount of waAjcr, 
it is necessary to generate a certain amount of heat by the combustion of fuel. The faeton 
controlling the amount of heat generated are: 

(a) Character of the fuel. 

(6) Intensity of draft. 

(c) Amount of grate surface. 

The cheapest fuel for the locality should be determined in advance by ascertaining the 
relative evaporating power of the coals available and their cost deUvered to the plant. 

Let G = area of grate surface, sq. ft. 

Wi = total equivalent evaporation from and at 212° F. 
W = equivalent evaporation per hour per lb. of fuel. 

"" B- 
R — rate of combustion, lb. fuel supplied grate per sq. ft. per hour. 

B _ 34.5 X b.hp. 
^ "i2 ~ WXR • 

With a good coal low in ash, approximately equal results may be obtained with a large grate 
surface and light draft or with small grate surface and strong draft, the total amount of coal 
per hour being the same in both cases. All fuels have, however, a maximum rate of combustion 
beyond which satisfactory results cannot be obtained regardless of the draft avaUable. 

With a coal high in ash, especially if the ashes are easily fusible, tending to chc^e the grate 
surfaces, a slow rate of combustion is required, unless shaking or travelling stokers are pro- 
vided to get rid of the ash as fast as formed. 

Types of Grate Bars. (Fig. 4.) The '^common grate'' is used for both wood and coal, but 
has been largely superseded by the " tupper or herring-bone grate" which is ordinarily furnished 
by the boiler manufacturer unless otherwise specified. 

The ''sawdust grate" is used only for sawdust as produced by sawmills and the "shaving 
grate" is used for burning shavings as produced by planning mills, sash and door factories. 

When bituminous coal is to be used, the front portion of the grate is frequently made solid 
for a depth of 6 to 12 inches, this portion being termed the ** dead plate," the purpose of which 
is to hold the fuel until the volatile products have been distiUed off. As soon as the charge is 
coked it>is pushed back and spread over the grate and a new charge introduced. 

The length of grate for bmning bituminous coals should not exceed about 6 feet for hand- 
fired boilers. If the grate has a greater width than 4 feet, two fire doors should be provided. 

Rocking Grates. The clinker formed on the grate, when bituminous coal is used, is more 
readily broken up and removed when a rocking or shaking type of grate is used, and the labor 
of stoking the fire is very materially reduced. The grate is ordinarily divided into two sectiona 
which permits of the live fire being shoved from one side to the other during the cleaning periods. 



BOILERS AND Ht'LES FOR CONsTRUCTlOiV 



85 



Fig. 5 shows A oommon type of rocking grate which is built in multiples of 6 inches in width 
and length. 

Ratio of Heating Surface to Gfate Area. The amount of grate surface required for a given 
condition with coal used as fuel will depend upon the rate of combustion assimied, which in turn 
is dependent upon the available cbaft, the quality and size of coal to be burned. 

The draft required at the boiler damper to produce various rates of combustion, and the draft 
between the ash pit and furnace is given by the curves, Fig. 3-D. The maximum grate area is 
limited by the design of the boiler for water tube boilers, manufacturers ordinarily providing 
for approximat^y 1 sq. ft. of grate surface for each 45 sq. ft. of heating surface. 







1^ ^ ^> 1^ 



£df 



COMMON GRATE 




>JJI»»»»lt^» 




TUPPER GRATE 




SHAVING GRATE 



I I I I if I J -I I J If I f I I -I .|..| .f I 



."■ .i! 



II' 



I 



SAWDUST GRATE 
Fio. 4. Types op Obatb Babb. 



TABLE 4 
AIR SPACES AND THICKNESS OF GRATE BARS 



Siie and Kind of Coal 



, Width of 
Air 



Ai 

Paaornot 

StOTO. . . . 

LAmp 
in' 
Wood: 




ThidoiMi 

of Onte 

Ban, 

In. 



H 
H 



88 POWER PLANTS AND REFRIGERATION 

When anthracite buckwheat is to be used a ratio of heatiog surface to grate area of I (o 
35 or 40 will ordinarily devebp the rated capacity <rf the boiler. 

When finer sizee are used or overloads are to be canied fenced draft must be employed to 
insure the desired results. 




RocKiNO Grate. 



The following tal)lo in ffven by GthhanWt " Sl<-am Power Plant Enuinpiring " as roprcarai 
IK current practice in this respect. 

TABLE 5 

RATIO OF HEATING SURFACE TO GRATE SDRPACE IN RECENT BOILER INSTALLATIONS 






''iiS^ 



IMcht 
Cbinmey 



ChmctaM FtM 



Cedtnlauiiona. , 



at <£ Wileoz Rntt 



The reader is also ref<Tred to the table of tests which gives the ratio of heatinn surface to 
grate area for water tube boilers, and the rate of combustioti and evaporation as obtained by thp 
use of various coab. 

The aecompanying table of relative values of steam coals ia taken from Meyet't "Steam 
Power Planla." The relative evaporative power of the better grades of a number of different 
coals is shown in the table, Pocahontas coal being placed at 100. These figures are approximate 
and should be used with some caution. The relative evaporation for the different coals shows 
what might be expected from the better grades of each kind of coal mentioned when fired by a 
good fireman under ordinary every-day conditions. 

Furaace Volume. Modem practice in the design of boiler furnaces is tending toward larger 
volumes and high settings, the object being to secure complete combustion before the products 
of combustion have reached the heating surface. 



BOILERS AND Rl(LES FOR CX)NSTRnCT10N 



KlndafCMl 




WUEvponl. s<^?S^ 

Condition. 


RMloof 






SB IB 411 




SO 
so 

s? 


5 
5 






t^S^JX"^'" 










S S 1 3B 





















Hi^ furnace temperaturea may be raaiBtaincd which allow the use of cheap eoala ruiminR 
hi^ in volatile matter without smoke. 

I/nr gas vdocitiee in the fimuce with comparatively hi|^ rat<ta of combuation are [MMtsible 
and fine conl may be burned without a large percentafte of it bointt cturicd over thf> briUite wall 



Fra. e. SaraoN or Watxh-Tubb Bon^n with Hioh Simvo. 

or to the tubes and stack. Tbe extra cost involving the use o1 high boiler setliiigs b maay tiotea 
ctbet by tb.a gain due to th^ use, and high settings are just as advantageous in small as in large 
plaiitB. 



88 POWER PLANTS AND REFRIGERATION 

At the Boston Elevated Railways Co. South Boston Station, the B, & W. boflers are 8 fL 
above the stoker grate at the rear. 

At the 20l8t Street Station, U. E, L, A P, Co., New York City, they are 10 ft. above. 

At the Dehuy plant of Detroit Edison Co. the gases travel 28 ft. from the underfeed stokefs 
before they strike the heating surface. An 84 in. by 18 ft. R. T. boiler in Newark is 72 in. above 
the grate. The lower row of tubes in~the double stoker boiler at the 59th Street power house of 
the Interborough Rapid Transit Co. are about 12 ft. from the stoker set under the rear of the 
boiler. 

Furnace Design for Smokeless Combustion. The following matter in reference to this 
subject is an extract from a paper before the Ohio Soc. of Mech., Elec and Steam Engineers, 1915, 
by Osbom Monnett, 

Conventional settings may often be changed at sli^t trouble and expense to give great 
improvement from a smoke standpoint, where high volatile, long flaming coals arc used. Hie 
type of boiler or furnace has less bearing on smoke performance than putting the combinatioo 
together so that both have a chance to give the best results. 

High Pressure Power Boilers with Chain Grate. Fig. 7-A shows in outline an old type, chain- 
grate setting with a 3H-ft. ignition arch, the stoker being set under the boiler with a clearance 
of 6 ft. from floor to front header. This setting is t3rpical of the older practice in chain-grate 
setting, with low, short, flat arch, poor ignition and low capacity. The deadening effect of the 
bank of tubes is such as to extinguish the flame before combustion has become complete, in the 
same manner that a wire netting will kill the flame from a gas burner, the result being a great 
deal of smoke. While this setting gives short flame travel, mere length of flame travel b not 
always enough to insure a satisfactory setting, imless some positive means are provided to cause 
a mixture, the gases frequently become stratified, in which case combustion cannot be complete 

In Fig. 7-B, the boiler has been raised to 10 ft. under the header; the ignition arch lengthened 
to 5 ft. and set full extension, which allows more flame travel, but the setting still has some of 
the defects of the first one and is not good for high capacities. One ot the principal defects is 
that the flow of rich volatile matter may pass into the bank of tubes in an uninterrupted current 
in the front part of the furnace, while most of the oxygen necessary to bum this volatile matter is 
passing in at the back part. There is a lack of mixture and consequently incomplete combus- 
tion and low economy. 

Fig. 7-C corrects the above defects by using a longer arch, setting the stoker farther under 
the boiler, decreasing the floor space occupied and narrowing up the furnace throat opening so 
that the volatile gases and air mix in a high temperature zone, which easily completes combustion 
on a 10-ft. setting. Experiments have shown that for commercial use the best throat opening 
is from 18 to 36 in., the smaller ones being high in maintenance; 30 in. is about the most satis- 
factory for all-around use. 

Another factor, which has had a marked effect on the performance of the later chain-grate 
settings, has been the height of the ignition arch at the grate; where 1 1 in. was formerly the stand- 
ard height for a flat arch, it has now been increased to 15 in., and the slope of the arch has been 
increased to 2 in. or 3 in. per ft. Where the arch is sprung across the. furnace, it is now set level, 
9 in. above the grate at the skewback, with a 9-in. spring, making 18 in. at the center of the 
arch. 

For the horizontal baffle little need be said from the smoke standpoint, as this combination 
is always satisfactory. Fig. 7-D shows a setting with 7 ft. 6 in. head room, which can be con- 
sidered ideal for a chain grate. This dimension may vary considerably without affecting the 
performance. 6 ft. 6 in. may be considered the minimum head room allowable. 

It sometimes happens that, with a tile-roof furnace and a low setting, the furnace gets so 
hot as to have a bad effect on the life of the brickwork. This can be offset in many instances 
by baring the lower row of tubes, using T-tile instead of box tile. This allows more rapid heat 
absorption into the boiler, increasing the life of the brickwork and resulting in a better operating 
furnace. 



BOILERS AND RULES FX)R CONSTRUCTION 




90 



POWER PLANTS AND REFRIGERATION 



Doublf-Inclined Slokat. Fck" the double inclined type o( atdker the BluHt length of flame, 
dischorgiiiK directly into the bank of tubee, is undesirable ythea the fire is being worked. This 
type of setting is frequently found inatalled in a 7-ft. head room, aa in Pig. &-A. Tbie huniaii 
element ent«rs strongly into the matt^" with euch a setting, owing to the poeaibility <rf having 
con^derable volatile matter pass off rapidly through carelessness. With a case of this kind it 



i 
A-poon 




in better to set the boiler with a clearance of 10 ft as in A, giving more oppcMlunity for the gase« 
to complete their combustion. One of the safest arrangemente is to provide a tile-roof setting 
with an auxiliary bridge well (Fig. 8-C), breaking up the current of gases and insuring the mixtun 
of any eicesa amount of volatile matter which may paae off for any cause whatever. The im- 
)>ortance of setting this type of furnace with maximum flame travel is not always realised. 

In Fig. 9 different types of boilers are shown with good and bad combination of double in- 
clined furnaces. It is a safe rule to get a full extension on this type of furnace and never resort 
to the flush front setting. In the case of Fig. ft-A, the defect o! short flame travel is correct«d 
by providing a 5-ft. dog-house extension between the boiler and furnace and by roisiog the boiler 
to^t the full benefit of the heating surface as shown in B. Topical Slirling settings are shown 
in C and D with flush front and full extension furnaces. 

Fronl-Feed Stokers. With the froat-feed stoker the same practice should be obso'ved aa 
regards flame travel. A clearance of 7 ft. is not sufficient to get good results with this type of 
HlokcT and vertically baffled water-tube boilers. A much improved furnace can be obtained by 
iixing a head room of 10 ft. as in Fig. 10-S, a combination reeulting satisfactorily frcHii every 
Ktiindpoint. This design alao gives an opportunity for employing a vertical bridge wall, wtudt 
is nearly always found to be a desirable feature wherever it can be used, as the radiating surface 
of the hot brick helpa to keep the gaaes hot as they pass out of the furnace. 

With n horizontal biLffle it is a simple matter to combine this type of stoker successfully, 
tjuflicient head room only is required to g^t the st«ker under the fwnl header. If this cannot bo 



BOILERS AND RULES FOR CONSTRUCTION 



91 



secured in the head room available, it does not alter the effectivenees of the design to excavate 
as shown in Fig. 10-0. Sometimes piers, or deflection arches, are used with this setting to break 
up the current of gases. Where a free opening in such a setting does not go below 40 per cent 
oC the grate surface of the stoker, such construction is desirable. On a vertical boiler always get 
the maximum extension possible ^thin reason. 

Underfeed Stokers. Difif^rent types require different head rooms (Fig. 11). The Jones and 
American types can give excellent results with a head room of 8 ft. 6 in. for a vertically baffled 



rticil BsfH$ 




A-POOR 



B-GOOD 




C-QOOD 0-6000 

Fio. 10. Front-Feed Stokers with Various Boilers. 



boiler, Fig. ll-B, and 7 ft. for a horizontally baffled boiler. In the case of the former the effort 
should be to provide enough flame travel to minimize the danger of unconsumed volatile matter 
passing into the bank of tubes. 

In the case of tubular boilers the above named types of stokers can be installed with 42 in. 
from the dead plate to the shell. Fig. 11-C, and the combination will result in a satisfactory 
perfonnance. With stokers of the inclined type. Fig. 11-D, a 10-ft. clearance under the front 
header makes on ideal combination. 

Hand-Fired Settings, One of the most common types of boiler setting encountered is the 
oidinary hand-fired, return-tubular setting such as is indicated in Fig. 12-^1 . In this setting there 
is no attempt made to accomplish a mixture of the gases after they have passed the bridge wall. 
The setting, while fairly efficient commercially, is very smoky with high volatile coal, and many 
attempts have been made to improve it. Fig. 12-B shows a full-extension, Dutch-oven setting 
by which it was att^npted to improve the plain hand-firing set ting. From a smoke standpoint 




Hamd-Fired Furnace. 



KOILEKa AND RULES tX)R CONSTRUCTION 



!)3 

the Dutcb-oven setting is a poor combinatioa. Cwttrary to stoker practice, where the fuel ia 
introdueed alowly and in aniaU quantities, thwe is & conaidemble quantity o( coal thrown on the 
fire at onoe. The strong radiatioa frcun the brickwork above the fire faas the effect of diatilling 
B so rapidly that pntTs of dense smoke will be made after firing in spit« of every effort 




RCTUPN TUBULAR 




Fio 13 Down Draft lEmvoa for Hkativo Load*. 

to prevent th«n. Fig. 12-C shows how to correct this defect by erpoeing the eheQ to the direct 
radiaticMi «rf the fire. Hue increases the steammg cqtacity and provides a high temperature 
■Mie back of the bridge wall where the gases must nix positively against the deflection arch, 
which breaks up the stratification and so promotes combustion. 

It is not practical to cmnbine a hand-fired, coal-burning furnace with a vertically baffled 
WBt<r-tube boiler, but it is a simple matter to arrange such a furnace with a horizontal baffle, 
carrying out the same idea as in Fig. 12-C. The ordinary band-fired, horizontally baffled water- 
tube boiler furnace is covered with box tile and has nearly all the defects of the Dutch oven shown 
in Fig. 12-B, as it is practically a fire-brick encloeed furnace from which the volatile gases will 
be distilled at a rapid rate. fig. 12-D indicates how this can be overcome. The changes in- 
dicated are, first, baring the tubes over the fire, using T-tile, thereby avoiding the radiating 
(ffeet of a mess of firebrick; second, instaUing a 2-Bpan deflectbn arch to break up the current 
of gases, as in the case of the return tubular boiler. In both of these furnaces a few simple pro- 
pmiiiKiB ^MMild be carried out to insure satisfactory results. 



04 ' POWER PLANTS AND RRFRICERATION 

There should be from 20 to 25 per cent of the grate surface in free opening above the bridge 
wall. The free opening from the back of the bridge wall to the d^cction arch should not be 
less than 40 per cent of the grate surface, while the free opening under the deflection arch shoukl 
be 50 per cent of the grate surface. Hand-fired furnaces for high-pressure work should be fitted 
with four air-siphon 8t«am jets, spaced across the furnace above the fire-doors, to be used when 
necessary. 

LauhPressure Healing Planla, The foregoing discussion has been with reference to hig^ 
pressure power work. The low-pressure heating plant presents a problem that in some respects 
is more difficult than any encountered in high-pressure work. The plants are not ordinarily 
large enough to justify stokers, and, even if such were the case, the character of the attendance 
is not such as would do justice to the equipment. The temperatures are lower and no steam 
is available for steam jets or for power to drive apparatus. With such conditions as these to 
meet it has been found that the down-draft principle works out very well. 

A Uttle study will show why this is so. The danger of making smoke on a down-draft 
furnace comes from getting green coal on the lower grate, so the longer the fire can remain un- 
disturbed the less chance of making smoke. The rate of combustion on heating loads is low, 
and allows for long periods during which the fires are not disturbed and no smoke is made. During 
these imdisturbed periods there is accumulating on the water grate a thick bed of coked ooal, 
which, when sUced down to the lower grate, does not make smoke because all volatile matter 
has been distilled off. After slicing, the fire can be heavily charged with fresh coal, without 
disturbing the fuel bed, consequently without causing smoke. It is then in shape for another 
long undisturbed period. 

Another advantage of the down-draft principle on heating loads comes from the fact that 
although the rate of combustion may be at times extremely low, yet the water element directly 
in the fire furnishes a proportionate amount of steam no matter how low the combustion; so the 
sjrstem is more responsive than would be possible with a plain grate boiler. 

The down-draft principle can be applied to return tubular or water-tube boilers in the larger 
units. In these units it is advisable to spring an arch in the path of the gas as shown in Figs. 
IS- A an4 13-B. As the rate of combustion on these large units at times approximates power 
conditions, it is desirable to guard against any excessive amount of volatile matter, which mig^t 
pass over during these periods, by breaking up the current of gases and giving them an oppm^ 
tunity to bum. 

For small units theie has been developed in the past few years a number of types of self- 
oont^ned, steel and cast-iron boilers embodying the down-draft principle. In the former type, 
Fig. 13-C, the water element consists of water tubes or pipes extended into headers in the ordi- 
nary manner and located in the firebox of a locomotive-type boiler. In the cast-iron, down-draft 
type. Fig. 13-D, the water element is cast integrally with each section, forming the upper grate, 
the shape of the elements being such as to f acihtate the slicing of coked coal down to the lower grate 
without disturbing the main body of fuel before the volatile matter has been distilled from it. 
This type is made in sizes up to 10,000 sq. ft. of radiation in one unit, and can be installed several 
in a battery. 

Design of Ash Pits and Hoppers. The capacity of the ash pit should be such as will ac- 
commodate the ashes accumulated during 14 to 16 hours' operation at maximum rating in order 
to avoid the necessity of. a night shift for ash handling. The weight and volume of ashes to be 
provided for may be approximated by assuming the boiler to be operated at 150 per o^t over- 
load and applying the following formula: 

Let b.hp. » normal boiler rating, boiler horsepower. 
C = calorific value of fuel. 

E = over-all efficiency of boiler grate and furnace. 
W = weight of fuel burned per, hour. 
a * proportion of ash in fuol. 



BOILERS AND RULES FOR CONSTRUCTION 95 

B 0.10 for high-grade anthracite. 
0.15 for Pittsburgh bituminouB. 
0.20 for Illinoifi and Indiana bituminous. 
0.40 for Iowa and eome southwestern localities. 
A -> weight (A ash per hour (1 cu. ft. ash « 40 to 50 lb.). 

-_ 1.5 X 33,524 X bJip. X E 

W ^ c • 

A ^aW. 

Szample. Required the capacity of ash hopper for a 300 hp. bofler for the following conditions 
of operation. Maadmum overload 150 per cent, calorific value of fuel 13,500 B.tu., 15 per cent ash, 
over-all efficiency 65 per cent. Hopper to hold the ashes for 16 hours' operation. 

^ 1.5 X 33.524 X 300 X 0.65 ^^^ .^ 

W — -— « 726 lb. coal per hour. 

13,600 

A » 0.15 X 726 - 110 lb. ash per hour. 

The volume to be provided is therefore; 

110 X 16 



40 



44 cu. ft. 



Topical Designs. In small plants in which not over four boilers (1000 biip. or less) are 
operated the ashes are ordinarily removed by hand. The most satisfacUMy type of pit, in this 
case, is a plain pit of rectangular section as shown by Fig. 14-il. 

The ashes must be hoed forward and then shoveled out. The distance the ashes are drawn 
should be limited to approximately 8 feet and there should be clearance for the hoe handle as the 
ashes are dragged forward and ample room for shoveling. 

In larger-eise plants provided with stokers a more elaborate system of ash removal is ordi- 
narily employed. Figs. 14-B to G and the description following were taken from ** Power/' 
Dec., 1914. 

Fig. I4r-B represents a sloping pit. Such pits are used where it is impossible to excavate 
to a pfoper depth, but the results are usually unsatisfactory. The design is a failure from the 
capadty standpcunt and also largely from the point of ash removal. 

Fig. 14-C indicates an ash hopper. These can be made of various designs and are frequently 
used in the largest stations. It is preferable to have such hoppers lined with fire-brick. They 
should have large valves, preferably 24 or 30 in. square, as this is a type in which the ash clinkers 
against the discharge vahre. Some designs have been made with diverging sides so that the ash 
and clinker cannot lodge. This pit indicates that it can be made of ample capacity and that if 
care be taken in the design the removal of ash is not difficult, but its upkeep is against it; in 
addition, it is difficult to inspect and to repair. 

The pit shown in Fig. 14-D is a desirable design and can be made to meet the proper re- 
quirements better than any of the others discussed. Ample capacity is provided and the ashes 
aft retained on a horis<Mital brick-lined or concrete floor and are not in contact with the metal 
discharge door. There is no tendency, therefore, for the door to warp and become leaky, and the 
ash itself remains in the same finely divided state in which it was discharged from the grate. 

Adies are removed by a hoe from this pit, and the designer ^ould make sure that the hori- 
lontal distance of the ash-pit floor is not over 8 ft. He should also see that there* is fully 8 ft. 
clearance in front <ji the door for the handle of the hoe when withdrawn. These pits when 6 ft. 
wide or under should have one 24-in. square cast-iron door; when in excess of this width two 
doors of this sise should be provided. 

This design has the three features of capacity, ease of ash removal and low maintenance. 



96 POWPJt PIJVNTS AND UK III! DERATION 

It can be modified to permit shoveling (rum the floor ia cases where the head room is slight ( 
modified to permit hoeing into a railway car where basemonta are designed along audi lines. 



I 



I 



Sometimes, due to soil conditions, to pipe lines or to other local conditions, it ia undesirable 
to excavate for an Bah pit in front of the bridge wall and the pit has to be put forward in front of 
the boilers. This necessitates the use of an ash drag or conveyor, as shown by Fix- 14-£ The 



BOILERS AND RULES FOR CONSTRUCTION 



97 



mechanism for such a device is simple and the results obtained are satisfactory. The arrange- 
ment costs niore than the plain pit, Fig. 14r-A, but requires less labor for ash removal, although 
this is partially offset by the additional labor consequent to an additional piece of conveying 
machinery. The c^>acity of such a pit is low, but the danger of overfilling it is removed the 
ashes being drawn forward where they can do no harm. 

Special types of ash-conveying machinery improve some of the designs materially, for in- 
stance, the plain pit shown in Fig. 14-A. When provided with a steam-jet system or a pneumatic 
ash-handling system, Fig. 14-f , it becomes a desirable design as to ease of ash removal, although 
the capacity of the pit is still limited 

The hopper arrangement, Fig. 14-C, lends itself to almost any type of conveying machinery 
or car system, as does that shown in Fig. 14-i>. 

The design shown in Fig. 14-£ can be improved by using the pneumatic ash-handling system 
or a steam-jet conveyor. Fig. 14-G. 

The designs shown by Figs. 15, 16 and 17 appeared in the " Practical Engineer," July, 1916, 
by R. A. Ixmgworthyf and refer particularly to ash removal by means of industrial cars run in 




Fig. 15. AsB HoppEB Under Room wrra No Basement — Side Dischasoe to Ash 

Tunnel in Front. 



the basement of boiler house. This method is perhaps the most practical and certainly an ex- 
tremely satisfactory method of ash removal whenever the layout may be adapted to this scheme. 

" The basemaiit under the boiler plant is most important and should be omitted only after 
a careful study has developed some particularly good reason for dispensing with it." The head 
room should never be less than twelve feet, and fifteen feet is preferable. 

Two good {ormB are shown by Figs. 15 and 16. The hoppers are illustrated merely to in- 
^cate types and are susceptible of modification to suit the plant under consideration. The 
hoppers should alwa3rs be of as large capacity as possible, so that they need not be emptied every 
time the fire is cleaned. They may be of steel, lined with concrete or brick, or of reinforced 
coQcrete, as the designs shown lend themselves readily to either form of construction. 

Fig. 17 shows an excellent hopper of large capacity. Its depth will depend upon the avail- 
able bead room, and this should be utilized to a maximum. The duplex gates should not be less 
than 24 in. square and, if the hopper is wider than about 6 ft., two gates should be used. The figure 
ahowB a fuU basement with thin partition walls to keep the dust and dirt from the remainder of 
the space. A tunnel might be used with this scheme, to save excavating the rest of the basement 
space; but if this is done, especial care should be taken to provide adequate ventilation. With the 



98 



POWER PLANTS AND REFRIGERATION 



oonstruction of Fig. 17, little handling of the ashes is required. The car should be of large 
ity and is run under the gates, filled and wheeled out. 

Fig. 15 shows another good type of ash hopper. It may be used as shown, which requirea 
the earoavatbn of a tunnel in front of a single line of boilers, or of the space between the fronts 




ttDfli;:*!*. .>/«>« •■litv .Tsrx-m.tenacxtt.xtiar.Tiai'Vrfx: 



Fio. 16. Side Dischabob Hoppbb to Reab Ash Tunnbl 








FiQ. 17. Labob CAPAcrrr Ash Hoppbb fob Dbbp Basembnt; Ash Tumnbl Bnclosbd. 



of a double row, or may be hung from the steelwork in a full basement. The dotted lines indicate 
the construction when used with stoker equipment. With this form of hopper, the ashes must 
be raked into the car, so that tiumel oonstruction should be made wide enough to allow a man to 
work handily. This construction will call for less depth of excavation than that shown in Fig. 17, 
but requires that a man handle the ashes from the hopper into the car. 

The construction shown in Fig. 16 involves a tiumel at the rear of the boiler, or may be used 
as a suiqsended hopper in the basement where the boiler settings are carried on the first floor 
steelwork. In the latter case, it leaves the center of the basement free for boiler feed pumps, air 
ducts, blowers, piping, etc. Some work will be required to get the ashes into the car, but it 
consists mainly in using a bar to keep them flowing and is an easier proposition than Fig. 15. 
Fig. 16 may be \ised without the full basement by excavating only for the a^ tunnel and helper. 

iVom the various figures shown and the description given, some idea of the requirements of 



BOILERS AND RULES FOR CX)NSTRUCTION 99 

a good ash hopper may be gathered. Make the hopper large enough and all parts heavy. Never 
use sheet steel in direct contact with the ashes; cast iron is best. 

EconomicAl Loads. The most economical "capacity rating" at which to figure the amount 
<ji boiler heating surface required for any proposed installation is dependent upon a number of 
items other than the cost of fuel alone. The cost of the power required as a percentage of the 
total cost of production in many lines of manufacture is small. For this condition the fixed 
charges (interest, insurance, taxes, etc.) are not very serious items, convenience in operation 
and insurance against breakdown being of greater moment. On the other hand, in plants where 
the product manufactured and sold is * 'power," economy in its production is obviously essential. 
The difficulty frequently experienced in correctly estimating or predetenhining the load 
curve which a proposed plant must eventually carry often makes it impracticable to make any 
but approximate calculations. In any event the boiler plant must be designed to handle the 
maximum estimated load with a sufficient reserve capacity to insure against breakdown. The 
"per cent rating" at which the boilers actually in service are to be operated to secure the maxi- 
mmn economy may be approximated from the data given below. 

When boilers are to be run beyond 133}^ per cent rating particular attention must be given 
to the question of providing a sufficient intensity of draft to accomplish the result desired, and 
f(» high overloads (200 per cent or more) ordinarily requires the installation of mechanical stok^v. 
The following matter has been condensed from the publication ** Steam": 
In a broad sense, all loads may be grouped in three classes: 
1st. Approximately constant 24-hour load. 

2d. The steady 10 or 12-hoiu: load usually with a noonday period of no load. 
3d. The 24-hour variable load found in central station practice. 
The economical load at which the boiler may run will vary with those groups. 
In figuring on the boiler load or the per cent capacity rating at which the boilers should be 
operated for best economy the broader economy is to be considered. That is, against the boiler 
efficiency there is to be weighed the first cost of the plant returns on such investment, fuel cost, 
labor, supplies, repairs, deprec^ktion, taxes, insurance, etc. 

1st. Constant B4-h(mr Load, For this condition of operation the most economical load will 
probably be found between 25 and 50 per cent above the rated capacity of the boilers. 

2d. The Steady 10 or 12'hour Load, Either an approximately steady load or one with a 
peak where the boilers have been banked overnight, the capacity at which they may be run with 
the best economy, all things considered, will be found to be higher than for uniform 24-hour load 
conditions. This is due to original investment, that is, a given amount of capital can be made 
to earn a larger return through the higher overload. 

Due to difficulties encountered in attempting to continuously operate at high overloads, the 
probable economical rating for this class of service will Ue between 150 and 175 per cent of rating. 
3d. The 24'hour Variable Load, This is the class of load carried by the central power station. 
In general where the maximum peak loads occur but a few times a year the plant should be of 
such a size as to enable it to carry these peaks at the maximum possible overload on the boilers, 
sufficient margin being allowed for insurance against continuity of service. 

With the boilers operating at this maximum overload through the peaks a large sacrffice 
in boiler efficiency is allowable, provided that by such sacrifice the overload expected is secured. 
Some methods of handling a load of this nature are given below : 

Certain plant operating conditions make it advisable, from the standpoint of plant economy, 
to carry whatever load is on the plant at any time on only such boilers as will furnish the power 
required when operating at ratings of, say, 150 to 200 per cent. That is, all boilers which are 
in service are operated at such ratings at all times, the variation in load being taken care of by 
the number of boilers on the line. Banked boilers are cut in to take care of increasing loads and 
peaks and placed again on bank when the peak periods have passed. It is probable that this 
method of handling central station load is to-day the most generally used. 

Other conditions of operation make it advisable to carry the load on a definite number of 






100 



POWER PLANTS AND REFRIGERATION 



boiler units, operating these at slightly below their rated capacity during periods oC H^t or 
low loads and securing the overload capacity during peaks by operating the same boilers at high 
ratings. In this method there are no boil^s kept on banked fires, the spares being epares in 
every sense <ji the word. 

A third method of handling widely varying loads which is coming somewhat into vogue 
is that of considering the plant as divided, one part to take care of what may be considered iht 
constant plant load, the other to take care of the floating or variable load. With such a method 
that portion of the plant carrying the steady load is so propcntioned that the boilers may be oper- 
ated at the point of maximum efficiency, this point being raised to a maximum through the 
use of economizers and the general installation of any apparatus leading to such results. The 
variable load will be carried on the remaining boilers of the plant under either of the methods 
just given, that is, at the high ratings of all boilers in service and banking oth»B, or a variable 
capacity from all boilers in service. 

TABLE 7 

STOKERS. COAL AND KILOWATTS PER BOILER HORSEPOWER IN LARGE CONDENSINQ 

TURBINE PLANTS. 



Plant 



Delny-DeCroit Ediaon Co 

20Ut St.. New York City 

69th St.. Now York City 

Watenide No. 2 

L St. Boston Ediaon Co 

So. Boston— Boston E. R. R. Co 

N. W. Sta.. Chieaco. Comm. Edison Co. 

Waterside No. 1 

Marion, Jersey City, P. S. Corporation . 



Kind of Stoker 



Underfeed 

Underfeed 

Underfeed and chain grate 

Underfeed 

Underfe e d 

Underfeed 

Chain grate 

Underfeed 
Chiefly underfeed 



KfaidorCoal 



Bituminous 

80% hard. 20% soft 

80% hard, 20% soft 

Bituminous 

Bituminous 18% Tolatfle 

Bituminous 

80% hard. 20% soft 

Bituminous— 24% volatile 



Kw. psr B.hp. 



Nonn* EmefK. 
4.22 6.66 
4.7 
4.2 
4.2 
4 

8.76 
8.46 
8.8 
8 



Example. Let it be required to calculate the amount of boiler heating surface for the following 
condition of operation. Ten or twelve-hour load having a peak, continuity of service essential, steam 
pressure 125 lb. gage, feed-water temperature 175**. Total weight of dry saturated steam required 
for the peak load W * 32,971 lb. per hour. Factor of evaporation, F b 1.083. Equivalent evapora- 
tion TT X F — 35,708 lb. Assume that a sufficient draft will be provided to successfully bum the 
grade of coal to be used at a rate of combustion necessary for 135 per cent capacity rating during the 
peak load periods. 

This requires an equivalent evaporation of 3.45 X 1.35 or 4.66 lb. per sq. ft. of heating surface 
per hour. 

Then to carry the peak load when the boilers, in service, are operated at 135 per cent rated capacity 
will require a total of 35,708/4.66 or 7.662 sq. ft. of heating surface, which is the equivalent of 7.662/10 
or 766 boiler horsepower. 

Subdivisioii of Heating Surface. The subdivision of heating surface for a plant (A this 
size would probably lie between the selection of 3-380 or 4-260 normally rated horsepower boi^rs 
where one spare boiler is considered sufficient as insurance against breakdown. If two spares 
are thought to be advisable and are recommended to provide for a contingency of having one 
boiler out of commission while one is being cleaned then the installation would consist of 4-380 
or 5-260 horsepower units. 

The curve, Fig. 3, will be found convenient when making a comparative study of the im>- 
portions for the heating surf aoe, grate arQa, size of chinmey or induced draft fan required for a 
particular installation. 

TYPES OF POWER BOILERS 

No attempt will be made to give a description of the numerous types of power bculeni In uae. 
Boilers are classified in general as either of the fire-tube type or the water-tube type. In the fire- 



BOILERS AND RULES FOR CX)NSTRUCTION 101 

tube boiler the hot gases pass through the tubes and in the water-tube type around the tubes. 
In so far as effici^icy is conoemedy exhaustive tests have proven that there is no choice between 
th» waler-iube and fiie-tube types. 

Return Tubular Boilers. This type of fire-tube boiler is the most common for use in small 
and medium-size installations. The standard brick setting is shown by Fig. 18. The boiler 
consists simply of a steel plate shell the heads of which form the tube sheets and into which the 
tubes are expanded. The shell is supported either by lugs, riveted to the shell, resting on wall 
plates or suspended by hangers from double channels supported by I-beam columns. 
The grate bars rest on castings anchored to the front and bndge walls. 
The gases pass under the shell over the bridge wall to the i>&ar combustion chamber then 
throui^ the tubes to the front, out the smoke connection to the breeching, and thenee, to the 
chimney. It is necessary to stay the flat surfaces of the tube sheets above the tubes. 

The boilers are manufactured either with or without steam domes. The primary object of 
the dome being to provide a separating space in order to obtain dry steam. The dome is being 
discarded for high-pressure work, as the large opening required weakens the shell and adds to 
the cost. In place of the dome a perforated pipe termed a ''dry*' pipe is often provided to prevent 
canying water over with the steam when the steam nozzle is connected directly to the shelL 
This type of boiler is built in commercial sizes from t5 to 200 hp. and pressures up to 150 lb. 
It is cheaper than the water-tube type and economical when properly operated. It requires little 
overhead room and affords a large heating surface in a small space. It is not found practical, 
owing to the internal stress set up in the shell by the large difference in temperature existing be- 
tween the inside and outside surfaces, to use plates much over one-half inch in thickness, the 
capacity at sise for a certain pressure being dependent upon the diameter, which in turn is a func- 
tion of the thickness of the plates, limits the construction of this type to smaller size units than 
may be obtained in the water-tube type. The water circulation is not so rapid as in the wat^- 
tube type, and they are therefore not so well adapted for rapid forcing to meet the varying demands 
of a widely fluctuating load such as is found in central-station work. For most manufacturing 
plant loads this type is, however, suitable and the insurance rates are no higher than with the 
water-tube type. See Fig. 20 and Tables 9, 10 and 11. 

Smok^ess Boiler Settings. The following matter, referring principally to return tubular 
boilers, has been condensed from a paper on " Smoke Prevention " by 0«6am Monnell,* He 
states that in general where water-tube boilers are installed there is plenty of space, and as a 
result little difficulty is encountered in abating smoke. The return tubular boiler, however, is 
chosen largely because of the limited amount of space which is available, and for this reason con- 
siderable study and planning must be done in order to prevent the furnaces from forming smoke. 
The ratio of grate area to heating surface is the most important problem, and for good prac- 
tice, where Illinois coal is burned, this should be 1 to 35 to 1 to 45, with an average of about 1 to 
40 for return tubular boilers. The grate area is usually too small, and for this reason the fires 
must be worked too frequently, thus causing them to smoke often and frequently excessively. 
He recommends the installation of some form of rocking or shaking grate and that the area 
above the bridge wall be not less than 25 per cent of the grate surface, also that the combustion 
diamber be kept clean down to the floor line. 

The hei^t of the boiler above the grate is an important consideration, and while former 
practice was from 22 to 24 in., the standard adopted by the Smoke Inspection Department* for 
boilen of 60 to 72 in. in diameter is 36 in. between the grat^ and boiler. Many furnaces which 
■noke have not sufficient gas space back of the boiler, and this dimension should be taken into 
consideration when smokeless combustion is desired. 

Another defect in the boiler settings frequently encountered is the restricted opening from 
the boiler shell to the uptake to the breeching. The method now adopted for relieving this situa- 
tion is to cut off a portion of the end of the shell and increase the size of the uptake, thus making 
sufficient amake area to carry the gases off without a great amount of friction. He recommends 

that 25 per cent additional space over the area of the tubes be allowed in the uptake where smoke- 
«Cliki«o, DUiiola. 



102 



POWER PLANTS AND REFRIGERATION 



less combustion is desired, and that for proper combustion there should be a draft of 0J22 in. 
over the fire in hand-fired furnaces. 

Common faults encountered in the breeching to boilers are that they are too small, too long, 
have too many turns and sometimes dips. Breeching should be as short and direct as possible; 
the* ratio for the area of the breeching to the grate surface should be as 1 to 4}^. This allows 
for a speed of the gases of 25 ft. per second with the boiler overloaded. 

In the accompanying table are given the stack dimensions for various sizes of horizontal 
return tubular boilers as recommended for use in Chicago. These heights are greater than 
ordinarily used in order to secure capacity, the problem having been presented to the Smoke 
Department not only to eliminate smoke, but to keep up cai>acity at the same time. 

This table applies only wfien the boilers are connected to the stack by a straight run of 
breeching which has fully as much area as the stack and in which long, narrow cross-s^stions 
and sudden changes of sections or drop in breeching are avoided. 



TABLE 8 
STACK DIMENSIONS FOR H. R. T. BOILERS 



Siaeof 


• 


Number or Boilbrs on Stack 


BoUer 


Tubes 


One 


Two 


• ■ - 
Three 


Pour 


48x14 
54x16 
60x16 
66x18 
72x18 
78x20 


84SM-in. 
34 4-in. 
46 4-in. 
54 4-in. 
70 4-in. 
84 4-in. 


21Hx 90 
24Hx 96 
28Hxl00 . 
31 X 110 
35 X 120 
39Hxl30 


80 X 100 
34 Hz 106 
40 X 110 
43 U X 120 
49^x180 
66 X 140 


37 X 110 
42^x115 
49 X 120 
53Hxl30 
60^x140 
68 xl50 


42Hxl20 
49 xl25 
57 xlSO 
61 X 140 
70 xl50 
78^x160 



The ordinary Dutch-oven types have become obsolete in Chicago at the present time. Ex- 
perimenting with baffles has shown that the double-arch bridge wall gives excellent satisfaction 
for smokeless combustion. A steam jet is sometimes placed in the furnace front, which is recom- 
mended for use after firing only. 

Horizontal water-tube boiler settings with vertical baffles and ordinary hand-fired furnaces 
have given considerable trouble in producing smoke, and have therefore been ruled out in Chicago. 
When a setting of this construction is encountered, the baffles are changed to the horizontal posi- 
tion, using tee tiles over the fire and box tiles over the combustion chamber, this arrangement giv- 
ing the smokeless combustion desired. 

The most difficult proportion which is met is that of mixed fuel consisting of shavings and 
coal. It has been demonstrated that shavings require about half the draft that is necessary for 
coal, and where the two fuels are used together, proper regulation of draft is a most difficult 
proposition for the fireman. One of the most common mixed-fuel furnaces is the full extension 
dutch -oven with horizontal baffles, these being of the box-tile type enclosing the tubes over the 
fire. A furnace which has met with considerable success in the burning of refuse under water- 
tube boilers is of the hand-fired type, using box tile over the grates and mixing baffles in the 
combustion chamber. 

Fittings and Connections for Horizontal Return Tubular (H. R. T.) Boilers (Fig. 18). 

Feed Ldne, Fitted with check and stop valve. 

Bailer Lead. Fitted with stop and gate valve. 

BlotD-qff Pipe. Run from lowest part of boiler insulated through heating space, pipe not 
over 2J^ in. diameter. For pressures over 135 lb., 2 valves or a cock and a valve are required, 
all extra heavy fittings. Free expansion must be allowed through brick setting. 

Pitch of Boiler not less than 1 in. in 12 ft. of length. 

LonffUudinal Joints to be above fire line of setting. 



BOILERS AND RULES FOR CONSTRUCTION 103 

Bndetta to fit oirrature of shell, not more than 2 riveta on each bradcet to cmne in same 
lon^tndiiiBl line. From top to bottom rivets in lug not less than 12 in. 

Brass or steel bcMler bushings through front head, open at end, dischai^ 3/S tiie distanc* 
from the fnxit head to the rear head below lowest water level in direction of natural circulation. 

Fmibte Plug. In rear bead, 2 in. or more above upper row of tubes. 

Water Column. Connection 1 in. or larger, steam from top of shell, water tram in. or mon 



Fro. 18. Frmiras and CotniEcnoim ros Rvrnix TuBm,&B BoiLnn. 

below center line of boiler, water connection of brass. Lowest part of gage glass above fusible 
plug and lowest safe water-line. Three gage cocks within visible length of gage glass. 

St^tty Valve, direct full opening. Discharge pipe direct full opening with open drain. 

Stop Valne in each steam outlet cm boiler oozEle. Provide drains where water accumulates. 

Steam Gage, Connected to steam space hy siphon of sufficient size to fill gage ti^ vrith 
water. No valve allowed except cock with T or indicating valve handle in pipe near gage. Dial 
graduationa to read Ifj maximiun allowable pressure. 

The Scotch Marine Boiler. This is an internally fired boiler of the fire-tube type, a lon^- 
tudinal section of which ia shown by Fig. 19. It is self-contained in that it requires no brick setting. 

This type has been used to some extent in office building in several of the larger cities. It 
requires httle bead room, has a minimum radiation Iocs, no leakage of cold air through faulty 
bri^ setting and requires a comparatively small amount of space for a given capacity. 

The circulation ia not so positive as in other types. 

The sise of the internally fired fire-tube boiler ia not limited by the thickness of the shell, as 
the fire comes in direct contact only with the furnace shell, which is subjected to a compreasi«Hi 
streoB and is of relatively small diameter. Boileta (A this type for marine use have been built 
in units of 500 boiler horsepower and designed to carry a working pressure of 200 lb. per aq. in. 
Thu type tA boiler is relatively expensive, althoi^ the cost of the boiler is offset somewhat by 
the abaenee of a brick setting. See Hg. 21 and Table 12. 

Water-Tube Boilera. The demand for boilers of larger siie than ia possible with the return 
tubular type, greater overload capacity, and abihty (o respond quickly to sudden demands has 
led to the selection of the watei^tube type of boiler for practically all modem large st«am in- 



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r4T^iH 


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no 


iHr^ 


00 00 00 

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to 00 A 

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r4<Hr4 


SSIS 


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^4 


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SS 


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j09MnKI 


s 


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106 POWER PLANTS AND BEFRJGKRATION 

Among the advantages of the wateMube boiler mar be menUtmed safety, aec e wibility tor 
quick Bt«aauiig, and capacity. Tlie latter two itema ore due to the more efficient dreulatiaD 
found in water-tube btrilers ct modem design. Expoienoe b» proven that the rate of he*t 



tranafcT through & metal miriatw from hot gasee to water ia depmdent upcm the velodty of the 
water over the surface and when a r^id circulation is secuted an increaoe in c^iaoity ia the 
natural result. It ia not unoommon to cq>eiate water-tube boiloa at ISO to 200 per cent at tbeir 
rated capacity. 

Tbe Babcock & ^Hcoz BoUer. Hie type <^ B.itW. boiler that ia conunonly employed in 
power plante ia shown by the accompanying Fig. 23 and Fig. 23. 

The boiler ia made up of one, two, or three longitudinal druma, 36" to 42" diameter, depend- 
ing upon the siie of the unita, conitected at the front and rear with the inclined tubea by meaoa 
at vertical tubea expanded into the cast-iron or preeaed-eteel headen and the forged steel cnws 
box riveted to the shell. 

The tubea, usually 1 in. in diameter and 16 to 18 feet in length, are expanded into headen 
of ainuouB form, which diapose the tubea in a ataggeied poeition when assembled aa a otnspleto 
boiler. Opposite each tube end in the headen there is placed a handhole of aufficient nae to 
p«iiut clMning and renewal of a tobe. The opening in castr-iron headers are made elliptjeal, 
and are closed by inside fitting forged plates with a milled face. The i^Mningi in the header 
have a raised milled seat. The joints between plates and headen are made metal to metal 
without gasket, and the plate ia held in position by studs and forged steel bindera and nuts. 



BOILERS AND RULES FOR CX)NSTRUCTION 



109 



mkm attsMKjriSr^'^ '^^ 







^mmimmmmmm^mmmmm m w^mmm% 



Dry Pip*' 



:a 



fM</ ri/t^« Ort0 on —oh t/tfi 
Utoo oomblno^ $top onf/ 



If 

M 
II 

;i 
II 
II 
II 
II 
II 
II 
II 
II 
II 

N 
li 
II 



Intomol food pipo to 
bo Ifttod ond toootod 
to iult ipooMootlano 
for tteh boitu\ 



^Bottom Blow 

I't^ltitortiol pfpo 
i^bqttom biqm^fF^ 



IHJHllI 



•I 
M 
II 

11 

11 
11 
II 

II 
II 
II 
II 
II 

J. " 

•<: i».. 



ood ouAiHofy food 



Otoln oook to bo 
loe^tod ot iow 
ondofbollBt* 



yt 



FlO. 21. OONNlOnONB lOB IMTBRKAL PUBNACB BOILBBS. 



PROPORTIONS PER SQUARE FOOT OF GRATE AREA FOR INTERNAL FURNACE BOILERS 




yi squv* Inch Main ttMin H aQUV* inch 

H SQUV* tndi Main and auxiliary faad Vm aquara indi 

Vit aqqara ineh Surf aea blowK>ff H bottom blow arta 



of wndkaitack « Vi grata aurfaea 

Ana ow bridfa walla « */y grata aorfaea 

Leaat area through tobea « Vr grata aurfaoa 

Ratio boating aurfaea to grata area » 80 or 86 to I 



Ctit M to hava about 45 par cant daar opening unla w otharwiae diractod* 



TABLE 12 

SPECIFICATIONS OF STANDARD SCOTCH BOILERS 



Tgth of ahdU in 
of ahdUin... 

ol JMMidk b.. 
I of flralwii ilua. 
„ AkngthoC 

boKHao. in 

Nombar off tubaa 

Diam. * I'gth off tubaa. In. 

8taa €rf dooM* to. 

8tai €rf pop aa iaty ya lTO.. . 
Shaofataan cwtlit, in . . . 
8tai of dMek A atop valra . 
8taa of blow-off vaiva.. . . . 
DinmaCv of Btack, in in . . 
Lmgthof Btacktin ft.... 
Nnmbar of ataal in atack. . 

Lnfftb off gratea 

1 1^^ > • . • . 



10 



8te5i 



18s42 



16 



18x18120x20 



24 

16 

81 H 
8800 



20 



42s00 44x70 



21x46 
62 
2x46 



25 



48X78 
0-82 
7-16 

H 

24x60 

68 

2x60 
24x24 

86 
16 
86 

4700 



80 



64x78 
5-16 
7-16 
18-82 

26x60 
82 
2x60 




85 



60x108 
11-88 

7-16 

28x90 
42 
8x90 
80x80 

2 

i^ 

85 

16 

42 

7900 



40 



66x104 

H 
7-16 

[80x84 
50 

8x84 
84x84 

2 

|H 
V^ 

40 
16 
60 
10,000 



50 



66x128 

H 
7-16 
7-16 

80x108 
50 
8x1081 
84x84 

r 

40 
16 
60 
11,200 



60 



72x128 
7-16 

11^ 



84x108 




70 



78x124 
7-16 

11^ 

86x102 
75 

8x102 
86x86 

8 

1» 

28 
45 
14 

72 
14,700 



80 



78x14278x15478x166 



7-16 
1(^2 



75 
8x1201 



90 



7-16 
11^ 



86x120 86x182 86xlU 



86x86 86x86 



8 

4 

28 
45 
14 

78 
16,1001 



76 
8x1821 



28 
50 
14 
78 
17,1001 



100 



7-16 
11^ 



75 

8xlU 

86x86 

J« 

28 
50 

14 
84 
18,000 



F Utuiaa faMtada nar amoka doon and firama, fumaea front with flra and ad& doora, grataa, ilua plata, amoka bonnat 



pop aaiaty valva* oomblnation watar eolumn with 54noh ataam gaga and aipbon, glaai watar 
elwak and atop valvaa, blow-off valva and ataok, and guy wire (four timaa kngth of ataeic). 



110 POWER PLANTS AND REFMGERATION 

Cast-iron headers &ie Dot recommeiided for pressures exceeding ISO lb. per sq. inch, T^ 
pressed-flteel bead is equipped with circular outside handbole fittings. Boilers oC the kuigitudiiul 
drum type are BUspeoded front and rear from wrought-vteel supporting frames carried by cron 
channels attached to eteel columns. Thia allows for contraction and expansion of the parts 
without straining the boiler or the brick setting. 

The forged steel drumheads are provided with manholes aikd platce. 

The mud drum to which the header sections are attached at Uie lower end of the rear 



headera is made up of a forged steel box 7^ in. square and of such a length u to be coa- 
nect«d to all of the headers by means of short nipples. 

The mud drum is furnished with a handhole for cleaning and tapped for the blow-<rff con- 
nection. 

Pitti'ngs. Each bcaltr is provided with the following fittings as part of the standard equip- 
ment (Fig. 23): 

Blow-off connections and valves attached to the mud drum. 

Safety valves placed on noxilea on the steam drums. 



BOILERS AND RUU» FOR CONSTRUCTION 111 

A vater ocAiimii eonneoted to tlie front of the drum. 
A steam gage attached to the boiler front. 

Feed-water connectioaB and valves. A flanged stop and check Talve <A heavy pattern is 
« tt * c bred directly to each dnmihead, cloBing automatically in case of a nipture in the feed line. 
The Gxturea that are supplied with the boilers oonaist of: 
Dead plates and supports, the plates arranged for a fire-bricic lining. 
A full Bet (rf grate bars and beuers, the l^ter fitted with expansion sockets for side walls. 



Fro. 23. Prmsos and CoKNBCWOtis fob Watsb-Tobb Ttpb Boiler. 

Flame bridge plates with necessary fastenings, and special fire-brick for lining same. 

Bridge wall girder for hanging bridge wall with expansion sockets for side walls. 

A full set of access and cleaning doors through which all portions of the pressure parts may 
be reached. 

A swing damper and frame with damper operating rig. 

There are also supplied with each boiler a wrench for handhole nuts, a wat«rHlriven turbine 
tube cleaner, a set of fire tools, and a metal steam hose and cleaning pipe equipped with a ^>ecial 
oosile for blowing dust and soot from the tubes. 

The Heine BoSer. Fig. 24 shows a bngitudinol section through the Heine boiler. 

In this type ot water-tube boiler the steam drum and tubes are parallel with one another 
and inclined at an angle of about 22 degrees with the horisontal. The tubes are expanded into 
a single steel plate riveted header front and rear. 

Opposite each of the tube ends, in the headers, tiiere is placed a handhole with cover plate 



^ 



Pio. 24. Hdinr Boiixk vith Cbain-Gbate Stoker and Sitpbbbutbb. 



Fia. 25. Sbctiom or Fabkeb Dowk-Plow Boiucb wrra Sdperbe&teb. 



BOILERS AND RULES FOR CONSTRUCTION 



113 



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^ MS V«o uf (O le <d^b^<D b^t- aiTt- cSaSo 



5iSee9t-oo9io9M«oo|A'«0r4S 



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00 00 ^o^S S o^c^S^O|S e|^o 9 S IS t» 0| 



|g§l|§|i§|l§§lll§| 

•> B B • ^ B k ft m k t> ft B ft « ft^ft ft 



3olol9lo|b^^l,l4'^^l>^lolo^^ 
^^^^.^i!^3obbr^^^^d9«e^^ 

*s;* I:' ^* 'a:* ^* k* ^* *x* ^* 

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t-^t«At*A^-At-A^At-Afr-At-A 

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114 POWER PLANTS AND REFRIGERATION 

to allow for deaiiiiig and tube repl&cematt. Horizantal b&ffling is employed, sb shown in tiie 
figure. 

The mud drum in this type is loca(«d inside of the steam drum. The feed water enUra 
through the front steam drum head and is conveyed direct through the feed pipe to the mud 
drum in which a greater share of the sediment is collected and may be blown <A thiou^ the ui^Mr 



FlO. 26. SnmJHO WATXB-lttBB BOn.ES with SDmBBKATEB at MiDDLI PASS. 

blow-oS cock or valve. The circulation is down through the rear header, thenc« through 
the tubes to the front header and into the steam drum. A baffle is provided over the 
front header in the steam drum as shown to prevent an exceas of free moiature being carried 
out with the steam. The Kteler, Union and Edgemoor boilers are somewhat similar m 
cooatruction, 

(MA«r Cmruium Type* of boilers ai« shown by Figs. 26 to 29. 



8XCIION or A Oabiu. Vbbticai. Watxb- 

TUB! BOIIXR. 



POWER PLANTS AND REFRIGERATION 



Fra. 39. Thb Bdosuook Waixb-Tuh Boiixb wim BvrwtaauTtm. 



TABLE 14 
PRINCIPAL DIUBHSION8 OP HEINIi ffTANDARD SINGLB-DRDU VATER-TOBK BOIUB 



k 



m 



IT tMM 

41,aM 




\tn 



NotK— HMthic Birbv ud •Uppliii nlclrt u* b«^ ca ■ tuba Uafth of 16 iMt. 
Rnurk— DinwBiioB (6) h torlB (t tnb^ Ibom (S) I" for anry 1 ft. ndantlaB bi iMWtk of MbL 
Parian. TatH (l)-cn. Fv 14 Ft. Tuba (D- 186. PBr 1« Pt. Tube* (1) - t09. PBr 11 Ft. Tnta ai-KL 
PorlS Ft. Totn (10) - 41K- For U Ft. TvbM (10) - 46K- Par IS Ft. TnbM (10) - 41}!. Fn 11 Ft. 




no. SI. Danmom or Hbmb DouBia-DHCH Boilbbb. 



118 



POWER PLANTS AND REFRIGERATION 



r«- 



4 »« 5 »•< ^ >j-» 

I 




Fio. 32. , DnCBNBlONB OF THE Stiruno Boiler. (See Tmble ik.) 



TABLE 15 
PRINCIPAL DIMENSIONS OF HEINE STANDARD DOUBLE-DRUM WATER-TUBE BOILER 

All Dimensiona In Inches. 



109 
'116 
128 
180 
137 
144 
161 
158 
168 



■I 



B 



ill 



2.726 
3486 
2^1 
8,882 
8,608 
4,128 
8.802 
4.852 
4.000 
4.580 
4.190 
4.800 
4.890 
6,080 
4.610 
6,280 



oqI 



2t^ 

1^ 



55.900 
69,600 
68.400 
62.400 
70.600 
74.800 
78.600 
77,800 



189 H 171 
196 m78! 
189 H' 171 ; 
196 H 178: 



204 
211 
204 
211 



186 
193 
186 
198 



163 
170 
163 
170 

177 H 
184 M 
177 
184 



150 
157 
160 
157 
168 
175 
168 
175 



147 
154 
161 
168 



150 
143 
150 
150 
167 
150 
157 



8 



148 400 



9 10 



107 
100 
107 
108 
115 
108 
115 



72 
72 
82 

74 



60 
60 
60 
60 
60 
60 
60 
60 



12 


18 


16 


18 


15 H 


18 


15 


18 


15H 


18 


15H 


18 


16 


18 


15H 


18 


16 


18 



14 



16 



lOHlOf^ 



12 
12 



16 



16 



17 



18 



Stbam 

NOSKLB 



19^20 



18 
18 
18 
18 
18 
18 
18 
18 



16 
16 
16 
16 
16 
16 
16 
15 



3olta 




Nora: — Heating ■orteee and Bhipping weight are bued on a tube length of 16 feet. 
Remarlc — Dimeneion (8) is for 18 ft. tube. Increaae (8) 2" for every 2 ft. reduction in 

length of tube. 
ForT8 Ft. Tubes (D- 238. For 14 Ft. Tubes (1) - 185. For 16 Ft. Tubes (1) - 209. 

For 12 Ft. Tubes (1) - 161. 
For 18 Ft. Tubes (11)- 41 M. For 14 Ft. Tubes (11) - 45 K. For 16 >t. Tubes (11) - 

43 K. For 12 Ft. Tubes (11) - 47H- 



BOILEBS AND RULES FOR CONSTRUCTION 



119 



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Fig. 83. Dqcenbionb or the Stxblino Boilir. 

(S«e Table 16.) 






FlO. 34. DiMENBIONB OP THB PABKBB WATKB-TUBB BOXLBB. 



.t^ir 



I 



BOILERS AND RULES FOR CONSTRUCTION 
TABLE 17 

DIUENSIONS OF THE PARKSR WATER-TUBE DOWM-PLOW BOILES 



The above dimensiona are baeed on 17" side walk and 24" dividing walls. The laat fi 
ea are double-ended, and the length of grate is the total for two graUs. 



DIUENSIONS I 



TABLE 18 

F MANNING BOILERS (Sw Pii- Ml) 



Rfl B.Hp. ' '^sSoT 


^E 


S- 




Numb* 

or 

Tubn 


%^ 






E6 

i 


2 

>> 


it 


M 

i 

St 

sw 








































1 ■ 





Coct of Water-Tabe Boilera. The boilers listed in Table 19 following represent actual 
isstaUatione and hold fOT territory within a radius of approximately 250 miles from Chicago. 
The following data may be used for estimating the cost of brick settings: 
Allow $20.00 per M for material and labor for common brickwork and S60.00 per M for 
fire brick in place. For boilers requiring fire-tile add approximately 10 per cent to the fire-brick 
cost. To the total add approximately 15 per cent for contractor's profit and contingencies. 
The above figures should result in a cost of brickwork setting of about 12.00 per boiler horse- 
powei based on a rating of 10 sq. ft. per b.hp. 



122 



POWER PLANTS AND REFRIGERATION 



TABLE 19 

COST OF HORIZONTAL WATER-TUBE BOILERS INSTALLED, INCLUDING BRICK SETTING 



Mmnufacturer 



Boilor horiopow 

Working prci ire (gage^ 

Heating mirfsee per b.t > 

Grate Mirfsee per b.hp. square feet 

Diameter eteam outlet 

Shipping wdght per b.hp 

Price per b.hp. f.o.b. cars 

Price per b.hp. erected induding brickwork 
Dimensions of brickwork for two boilers 

erected in one battery: 

Front width 

Hdght 

Depth 



1 

A 


B 


C 


A 


C 


200 


200 


200 


800 


800 


150 


160 


150 


150 


150 


10.1 


10.36 


10.2 


10 


10 


0.214 


0.216 


0.220 


0.198 


0.226 


6" 


6" 


8" 


8" 


6" 


195 


210 


228 


228 


196 


$9.00 


$9.88 


fll.80 
$14.65 


$10.65 
$18.40 


$10.58 
$12.72 


$11.60 


$15.06 


21'-0" 


2S'-0" 


lO*- 6" 


24'-2" 


22'-9" 


16'-5" 


16'-8" 


16'- 8" 


18'-2" 


15'-1" 


16'--8" 


20'-2" 


17'-10" 


19'-9" 


20'-9" 



800 

160 

10. 3S 

0.217 

7" 

181 

$8.33 

$10.45 



28'-0'' 
20'-l" 
16'-y 



RULES FOR CONSTRUCTION AND INSTALLATION OF STEAM BOILERS 

FORMULATED BY THE BOILER CODE COMMITTEE OF THE 
AMERICAN SOCIETY OF MECHANICAL ENGINEERS* 

1. Mazhntun Allowable Pressure. The maximum pressure to be allowed on the shell (x* drum 
of a boiler shall be determined by the strength of the weakest course, due consideration being 
given in each course to the thickness and the tensile strength of the plate, the efficiency of the 
longitudinal joint, the inside diameter of the course and the factor of safety allowed by these 
rules. 

— ^-^ -— — — maximum allowable working pressure, per sq. in. in lb. 

T. S. » tensile strength of shell plates, in lb. per sq. in. 
t » minimum thickness of shell plates, in in. 
% ■> efficiency of longitudinal joint, method of determining which is given in 

Par. 74 of these rules. 
R » radius — one-half the inside diameter of the outside course of the sheU 
or drum, in in. 
F, S. ^ lowest factor of safety allowed by these rules. 

2. The steam pressure allowed on a boiler constructed entirely of cast iron, with the excep- 
tion of the connecting nipples, or a boiler built of steel plate to be used exclusively for low- 
pressure heating, shall not exceed 15 lb. per sq. in. This does not apply to economisers. 

3. The maximum pressure allowed on hot-water boilers for heating buildings or water for 
domestic purposes, constructed entirely of cast iron, with the exception of the connecting nipples, 
or a boiler built of steel< plate to be used exclusively for low-pressure heating, shall not exceed 30 
lb. when the temperature of the water in the boiler is less than 212 deg. Fahr., unless the boiler 
has been tested by a hydrostatic pressure to not less than twice the working pressure. When 
the water in the boiler is 212 deg. Fahr. or over, the maximum pressure allowed may be 30 lb. 
provided the boiler is tested as hereinafter specified. 

4. The pressure allowed on boilers foi heating purposes made wholly of cast iron, or on steel 
boilers with cast-iron mud rings, door frames and manhole flanges, fitted with an approved! type 
lock-pop safety valve, shall not exceed 15 lb. per sq. in.; and on all steel boilers with steel or 
wrought-iron mud rings, door frames and manhole flanges, shall be allowed a pressure not to 
exceed 50 lb. per sq. in. The steel, rivets and materials used in steel boilers, method of manu- 
facture, riveting, bracing, etc., iohe A. S, M, E, standard throughout, except that in no case 

* Extnet from Prngrwi Report Boiler Code Committee. Engineers ere now epedfying that boilen be con- 
•tmeted imder tbeee ruleSi 

t Spedfled by the State Boiler Department. 



BOILERS AND RULES FOR CONSTRUCTION 123 

ahftll steel of kse thsji iiia.us thicknesi, nor tube ^eets or heads of less than */u in- in thick- 
neee be used for any i^esBtire, and in no case shall the shell or drum of a steel boiler be used for 
heating purpoees provided for in this section, at a working pressure having a factor of safety id 
leas than t«n (10). Boilen provided Sot in this section shall be fitted witii approved safety 
devices. Each bailer must be provided with safety valve of the apring-pop type which cannot 
be adjuBted to a higher prcasure than 15 lb. per eq. in. for cast iron or partially cast-iion boilen 
and SO lb. per sq. in. for all-steel boilers, such valves to be of the lock-pop t)^ and bearing in- 
spector's tag showing compliance with state requirements. A certificate of inspection must be 
furnished with every boiler covered by this section, giving detailed description and thickness 
of metals used covering all the details of construction and test. 

5. The maximum pressure allowed on steel or wrought-iron water heaters, connected to 
open tank on roof or the local water supply system, shall not be over one-half the prCMure at 
which same was tested and marked by the manufacturer. Every such heater is to be tested 
St twice the allowable working pressure and fitted with manufacturers' braes name-plate marked 
plainly with the pressure at which vessel was tested and the allowable working pressure which 
is not to exceed one-half the said test pressure. 

6. The pressure allowed on a water-tube boiler, the tubes of which are secured to cast- or 
nulleBble^ron headers, or which have oast-iron mud drums, shall not exceed 160 lb. per eq. m. 

le header at any point shall 
gtb <rf a continuous cast or 
uction shall show a precBUre 
I lb. per sq. in. gage pressure 

x>nform with the following 

i from a good quality, dose- 





PwCcBt 




S.OO 

!l 

oioa 
















s hereafter mentioned shall 



coQiorm to tbe following Umite i 


a chemical compoeition: 






P«C«rt 




O.TO 









(c) PhytierU Propertiea. Two test bats 1 in. square on 12-in. centers shall show under a 
centrslly applied load an average tronsveise strength of 2800 lb. and an average deflection of 
not less than 0.14 in. 

(d) Two tensile test bare not less than 2 in. in length in the stressed sections shall show an 
average tensile strength of not less than 25,000 lb- per sq. in. 

8. like test bars referred to in (c) and (d) shall be prepared as foUows; 

On each lot oi not more than twenty headers cast at one heat in the foundry Uiere shall be cait ' 



124 



POWER PLANTS AND REFEUGERATION 



on two of the headers of the lot, or on the gate or sprue attached to the headers, a bar 1 in. square 
and 14 in. bng, which shall be used as the test bar in the test for transverse strength. When 
the two bars representing a lot <^ headers are brdLen, one of the pieces of each of the two bars 
shall be made into a test bar for tensile strength, with a diameter of Ji in. fcx* a length of 2 in. 
in its middle pcHlion. 

9. The pressure allowed on a boiler fitted with a state boiler department lock-pop safety 
valve shall not exceed 15 lb. per sq. in. This special type of safety valve is provided fw by Par. 1 
of the Engineers' and Firemen's License Law, and applies to boilers used for heating purpoees 
exclusively. 

10. Tensile Strengtfi. When the tensile strength of steel or wrou^t-iron shell plates is nU 
known, it shall be taken as 55,000 lb. for steel and 45,000 lb. for wrou^t iron. 

11. Crushing Strength of Mild SteeL The resistance to crushing of mild steel shall be taken 
at 95,000 lb. per sq. in. of cross-eectional area. 

12. Shearing Strength of Rivets. The shearing strength of rivets per sq. in. of croas-sectional 
area may be taken as follows, and these values were used in calculating the examples of joint 
efficiencies: 



Iron rlTets in dn^ shear 

Iron riTets in double shear t 

SMteel rivets in sin^e shear 

Steel rirets in double sheer 



88,000 
70,000 
42,000 
78,000 



13. The maximum shearing strength of rivets per sq. in. of cross-sectional area shall be: 



Iron rlTets in sinrie sheer. , 
Iron rivets in double sheer. 
Steel rivets in sin^e sheer. 
Steel rivets in double sheer 



Pounds 



88,000 
76,000 
44,000 
88.000 



14. Table 20 gives the allowable shearing strength of rivets from ^Vie in« to 1 ^/le in. in diame- 
ter, in lb., based on values given in Par. 12. 



TABLE 20 

ALLOWABLE SHEARING STRENGTH OF RIVETS 



Diemeter of rivet 














after driving, in... 


0'li8?5 


0.7^ 


.%l» 


0.876 


.%5 


1.0826 


Cross soetional area 














of rivet after driv- 














ing, sq. in 


0.8712 


0.4418 


0.6186 


0.6013 


0.8908 


0.8866 



Allowable Shearing Strength, Pound 



Iron, sinidei 
Iron, doable •«»». 
Ste«, sinrie sheer . 
Steel, double 



14,106 
26.984 
16,690 
28,964 



16,788 
80.926 
18,666 
84,460 



19.708 
86,296 
21,777 
40,448 



22.849 
42,091 
26,266 
46,901 



26,281 
48,821 
28,998 
68,848 



88,691 
62,068 
87,287 
69,166 



15. Rivets. When the diameter of the rivet holes in the longitudinal joints of a boiler is 
not known, the diameter and cross-sectional area of rivets shall be ascertained by cutting out 
one rivet in the body of the joint. 



BOILERS AND RULES FOR CONSTRUCTION 



125 



16. Fidon of Safety. Boilers in service one year after the passage of this act shall have 
a factor of safety of not less than four (4). 

Effective three (3) years from passage of act, the lowest factor shall be 4.25. 
Effective five (5) years from passage of act, the lowest factor shall be 4.50. 
Effective seven (7) years from passage of act, the lowest factor shall be 4.75. 
Effective nine (9) years from passage of act, the lowest fact<Mr shall be 5. 
Boilers built after passage of this act shall have a factor of safety <^ not less than five (5). 
These rifling factors do not apply to water-tube boilers in which the drums are not subject 
to the radiant heat of the furnace. 

17. No li4>-seam hmrisontal return tubular boiler over 36 in. in diameter carrying over 100- 
lb. pressure will be allowed in service after ten years from the passage of act. 

18. Age Limit. The age limit on lap-seam horisontal return tubular bmlers over 36 in. in 
diameter carrying over 50-lb. pressure shall be twenty (20) years. 

19. Siie of Safety Valves Not ^ring-Loaded. The minimum sise of safety valves (other 
than direct spring-loaded safety valves) shall be governed by the pressure allowed, as stated in 
the certificate of inspection, and by the grate area of the boiler, subject to the foUowing conditions 
and as shown by Table 21: 

(a) A single boiler, or two or more boilers connected to a common main and allowed the 
mtme prtMnare. The minimum sise of safety valves for each boiler shall be governed by the pres- 
sure aDowedy as stated in the certificate of inspection, and by the grate area of the boiler. 

(6) When two or more boilers, which are allowed differerU pressures, are connected to a com- 
mon steam main, the minimum siie of safety valves on each shall be governed by the pressure 
allowed, as stated in the certificate of inspection, and by the grate area of the boiler; and all 
safety valves shaU 6e set at a pressure not exceeding the lowest pressure allowed. The aggregate 
vahre area shall not be less than that required (or the aggregate grate area, based on the lowest 
pressure allowed, as shown by Table 21. 

(e) When two or XDore boilers, which are allowed differerU pressures, are connected to a com- 
mon steam main, and all safety valves are nolset&t & pressure not exceeding the lowest pressure 
allowed, the boiler or bcnlers allowed the lower pressures shall each be protected by an additional 
safety valve or valves placed on the connecting pipe to the steam main; the area or combined area 
of the safety valve or valves placed on the connecting pipe to the steam main shall not be less 
than the area of the connecting pipe, except when the steam main is smaller than the connecting 
pipe, when the area or combined area of safety valve or valves placed on the connecting pipe 
shaU not be less than the aiea of the steam main. Each safety valve placed on the connecting 
pipe shaU be set at a inessure not exceeding the pressure allowed on the boiler it protects. 

20. Table 21 gives the areas of grate surfaces, in square feet, for other than direct spring- 
loaded safety valves. 

TABLE 21 



SIZES OP SAFETY VALVES, NOT SPRING-LOADED, RELATIVE TO STEAM PRESSURE AND GRATE 

AREA 


Ifttinam Prtaan AOowad per 


Zero to 


Over 26 to 


Orer 60to 


Sqnai* Inch OB B«iter 


25 Pounds 


SOPounde 


100 Pound* 


DiuMtarolVahr^ 

iBCfaM 


AimolVmlT*, 
Sqnai* IndMi 


Atm 9t Grmte, Square F^et 


1 


0.7864 


1.50 


1.76 


2.00 


1^ 


1.2272 


2.26 


2.60 


8.00 


1.7S71 


8.00 


8.76 


4.00 


2 


8.1416 


5.60 


6.60 


7.26 


SH 


4.9087 


8.26 


10.00 


11.00 


8 


7.0686 


11.76 


14.26 


16.00 


SH 


9. 6211 


16.00 


19.50 


21.76 


4 


12.6660 


21.00 


25.60 


28.26 


4H 


16.9040 


26.76 


82.60 


86.00 


6 


19.6860 


82.76 


40.00 


44.go 

• 



126 POWER PLANTS AND REFRIGERATION 

21. Safety Valves. Each boiler shall have two or more safety valves, excepting a boiler 
that carries pressure not exceeding 15 lb. (gage) per sq. in., and excepting any boiler f<x> whidi 
one safety valve 3-in. size or smaller is required under Pars. 20 and 21 of this section. 

22. The discharge capacity of a spring-loaded pop safety valve shall be in accordance witli 
the values in Table 22. The discharge capacity of the safety valve, or i more than one safety 
valve is used on a boiler, the minimum aggregate discharge capacity of all of the safety valves, as 
shown in Table 22, shall be not less than the maximimi evaporative capacity of the boiler, based 
upon the mft^mum amount of fuel that can be burned in any one hour, and the heating value 
of the fuel, in accordance with Table 23, subject to the following conditions: 

(a) For a single boiler, or when two or more boilers are connected to a common steam main 
and allowed the same pressure, the minimum number and size of the safety valves required for 
each boiler and the minimum aggregate discharge capacity of all of the safety valves on each 
boiler shall be governed by the working pressure allowed, as stated in the certificate of inspection, 
and by the maximimi evaporative capacity of the boiler, calculated in accordance with Table 22. 

(&) When two or more boilers, which are allowed different pressures, and are carrying the 
same steam pressure are connected to a common steam main, the minimum number and size of 
the safety valves for each boiler shall be governed by the total evaporative capacity of the boiler, 
calculated in accordance with Table 22, and the lowest working pressure allowed upon any of the 
boilers, as stated in the certificate of inspection, and in accordance with the values in Table 22, 
and no safety valves shall be set at a pressure exceeding by more than 5 lb. the lowest working 
pressure allowed. The aggregate discharge capacity of all of the safety valves shall be at least 
sufficient to discharge, at the lowest pressing allowed on any of the boilers, the total maximum 
evaporative capacity of all of the boilers, in accordance with the table in Table 22. 

(c) When two or more boilers, which are allowed and are carrying different pressures, are 
connected to a common steam main, the minimum number and size of the safety valves on each 
boiler shall be governed by the maximum evaporative capacity of the boiler, calculated in accord- 
ance with Table 22, and the allowed working pressure for the boiler, as stated in the certificate 
of inspection, and in accordance with the values in Table 22, and to protect the low-pressure 
boilers thus connected additional safety valves shall be provided on the low-pressure piping. 
The aggregate discharge capacity of the additional safety valve or valves on the connecting pip- 
ing shall be at least sufficient to discharge, at the working pressure allowed on the boiler carry- 
ing the lower pressure and connected by such piping, the total evaporative capacity of the boiler 
or boilers carrying higher pressure; and no such additional safety valves on any connecting 
piping shall be set at a pressm^e exceeding by more than 5 lb. the working pressure allowed on 
the boiler connecte4 by such piping. 

23. A table oi discharge capacities foi direct spring-loaded pop safety valves follows. • The 
discharge capacity of a safety valve is expressed in equations 2 and 3 as the product of values 
C and H, The discharge capacities given in Table 22 are for each valve size at the pressures 
shown and are calculated upon the given values of lifts, which have been approved by safety 
valve manufactiu^rs. 

C = total weight, in lb., of fuel of any kind burned per hour at time of maximum forcing. 

(See Note, page 128.) 
H = the heat of combustion, in B.t.u. per lb. of fuel used. (See Note, page 128.) 
D = diameter of valve seat, in in. 
L « vertical lift of valve disc, in in., measured immediately after the sudden lift due to 

the pop. 

P = absolute boiler pressure per sq. in., or gage pressure plus 14.7 lb. 

The boiler efficiency is assumed as 75 per cent. 

Discharge efficiency of valve, based upon Napier^a formula, is taken as 96 per cent. 

CX H X 0.75 3.1416 XDXLX 0.707 X P X 0.96 ^ , . ,^ .. ^ 

= — — — . fQj. valve with 45-deff. seat. (1) 

1100 X 3600 70 ■»«««». »c»i;. },ij 



BOILERS AND RULES FOR CONSTRUCTION 127 

C H - 160,856 -KP XD-X.HOT valve with bevel seat at 45 deg. . . . (2) 
C H " 227,487 X P X /> X t ttw valve with flat seat at 90 deg . . . . (3) 
lU.1ubratio¥^». A boiler at the time of ina;iinnim forcing uses 2150 lb. of Blinoia (Marion 
County) coal per hour. Boiler preseure, 225-lb. gage. 

21S0 X 12,100 ~C H ■= 26,015,000 
This requires two 4-in. valves with 4&<leg. bevel seat, or one 4}^in. and one SJ^n. valve with 
4&^1^. bevel seat. 

Wood shAvingB of heat of combustion of 6400 B.t.u. per lb. are burned under a boiler at 
tiie maximum rate of 2000 lb. per hour. Boiler pressure 100-lb. gage. 

2000 X 6400 = C H = 12,800,000 
Thit requirea two 3^in. valves with 45-deg. bevel seat. 

12,800,000-i-IlOO - 11,637 lb. of steam discharged per hour. 
An oil-fired boiler nt maximum forcing uses 1000 lb. of crude oil (Texas) per hour. B<uler 
pressure 275-lb. gage. 

1000 X 18,500 -C H ~ 18,500,000 

TABLE 22 

DISCHARGE CAPACITIES TOR DIRBCT SPRING-LOADED POP .SAFETY VALVES 
DluBMar of Vain 



This requires one 3^in. and one 3-in. bevel-seated valve, or one S-in. and one 2}^in. valve 
with flat seats. 

18,500,000/1100 - 16,818 lb. of steam po' hour. 



128 POWER PLANTS AND REFRIGERATION 

A boiler fired with natural gas consumes 3000 cu. ft. per hour. The safety valves are set 
at 150-Ib. gage. 

3000 X 960 = CH = 2,880,000 
This calls for one 2-in. valve. « 

For waste-heat boilers C is the maximum weight of gases supplied to the boiler per hour 
and H X 0.75 should be replaced by Cp (d — (t) X 0.97, where: 

Cp B specific heat of the gases at constant pressure. 

ti « initial temperature of the gases in deg. Fahr. 

it B final temperature of the gases in deg. Fahr. 

For most waste-heat work sufficient accuracy is secured by determining the numerical value 
of the term 0.33 (tx — t%) and using this in place of H in the takl \ 

Illustration: Assume C » 90,000 lb. of gas per hoiu*; (i = 2000 deg. Fahr.; it ^ 450 d^ 
Fahr.; boiler pressure, 150-lb. gage. 

C X 0.33 (ii - it) = 46,035,000 

Three 4-in. valves would be required. 

Note, The heat of combustion if not known may be determined by a coal calorimeter. The 
report of the coal-testing plant of the United States Geological Survey made in 1904 shows the 
values for the heat of combustion for coals ordinarily used in the United States as given in Table 23. 

In the absence of more exact data, the values of H in B.t.u. per lb. may be assumed for 
various fuels, in accordance with the following: 

TABLE 23 



Semi-bitumiiioiu ooal 14,500 

Anthndte 18.700 

Seroeninei ! 12,500 

Coke , 7,800 

Wood, hard or •oft. Idln-dried 7.700 

Wood, hard or wott, air-dried 6,200 

Wood ■havinsi , 6,400 

Peat, air-dried. 25 per eent moisture 7,500 

Ufnite I 10.000 

Keroeene, per pound 20,000 

Petroleum, crude ofl, Penn 20,700 

Petroleum, crude ofl, Texas 18,500 



In determining the number and size of safety valves for a boiler using gaseous fuel,C becomes 
the cu. ft. per hour supplied at time of maximum forcing, and H the higher heating value p^ 
cu. ft. The higher heating value is used inasmuch as the boiler efificiency with a gaseous fuel 
is generally higher than the 75 per cent efficiency assumed in the formula. The values of H may 
be assumed, in B.t.u. per cu. ft., at 62 deg. Fahr., as follows: 



Natural gas 

Blast^fuinaee gas 

Producer gas 

Water gas, uncarburetted 



960 
100 
150^ 
290 



24. The discharge capacity of a safety valve shall be rated in accordance with the values in 
Table 22. For pressures intermediate between those given in the table, the number and sixes 
of the safety vidves required shall be determined at the nearer pressure shown in the table. The 
amount of the vertical lift of the valve disc from its seat, measured immediately after the sudden 
lift due to the pop, must be not less than the value given in Table 22 for the corresponding 
valve size and pressure. The lift is to be measured with the blow-down adjusted in accordance 
with Par. 32. The maximum lift of any safety valve for boilers must not exceed fifteen one- 
hundreths (0.15) in. The discharge rating of a safety valve shall not be greater, for any valve 
siae and pressure, than that given in Table 22. 



BOILERS AND RULES FOR CONSTRUCTION 129 

TABLE 24 

VALUES TOR THE HEAT OF COMBUSTION FOR COAIS ORDINARILY USED IN THE UNITED 



NaaootStata 


"■"-HisS" 


Carbon 

-fisr 


1i%r 


„_ ,„„,. 


uJiSRSSSd 

Wairior Field 

UcAI«««rBed 
Uarion County 

Polk County 

BdlavUlen^d 
ApMOOOae County 

|i 'l^ 

Bl laid 
Bl idd 








£0 
i 


1 

J 
i 

2 
E 
Z 
S 


































l.MZ 

























































































































































































25. SttTety valves hereafter insUUcd on boilers shall not exceed 4H in. nominal seat diameter, 
meamued at the inner edge of the valve seat; and no safety valve used on a boiler shall have the 
valve seat leas than 1 in. diameter. 

26. Safety valvea shall be the direct spring-loaded pop type, with seat and bearing surface 
of the disc either inclined at an angle of about 45 deg. or flat at an angle of about 90 deg. to the 
center line of the spindle. 

27. When two or more safety valves are used on a boiler, one valve shall be set to open at 
the allowed pressure stated in the certificate of inspection, and the Other valve or valves shall 
be set to open at pressures at least 3 lb. or 5 lb. higher. If all the valves on a boiler are not of 
the Bane edie, tiie valve that is set to open at the allowed pressure shall have a discharge capacity 



130 POWER PLANTS AND REFRIGERATION 

at least as great as the maximum evaporative capacity of the boiler divided by the number of 
safety valves on the boiler. 

28. When two or more safety valves are used on a boiler, they may be either separate valves 
or twin valves made by mounting separate valves on Y bases. 

29. The safety valve or valves must be connected to the boiler independent of any other 
steam connection, and attached directly to the boiler or as close as possible to the boiler, without 
any intervening pipe or other fitting between the valve and the boiler, except the Y base forming 
a part of a twin valve or the shortest possible nipple or bushing. A safety valve must not be 
connected to an internal pipe in the boiler. Every safety valve shall be connected so as to stand 
in an upright position, with spindle vertical, when possible. 

30. When a boiler is fitted with two or more safety valves on one connection, this connection 
to the boiler shall have a cross-sectioiial area not less than the combined area of all of the safety 
valves. 

31. Each safety valve shall have full-sized direct connection to the boiler. No valve of 
any description shall be placed between the safety valve and the boiler, nor on the discharge 
pipe between the safety valve and the atmosphere. When a discharge pipe is used, it shall be 
not less than the full size of the valve, and the discharge pipe shall be fitted with an open drain 
to prevent water lodging in the upper part of the safety valve or in the pipe. When an elbow is 
placed on a safety-valve discharge pipe, the elbow shall be located close to the safety-valve outlet 
or the pipe shall be securely anchored and supported. All safety-valve discharges shall be so 
located or piped a^ to be carried clear from running boards or working platforms used in control- 
ling main stop valves of boilers or steam headers. 

32. Safety valves shall be set and adjusted as follows: To close after blowing down at 
least 2 lb. on boilers carrying allowed pressure not exceeding 15-lb. gage; to close aft^ blowing 
down at least 3 lb. on boilers carrying pressures over 15-lb. gage and up to and including 125-Ib. 
gage; to close after blowing down at least 4 lb. on boilers carrying over 125-lb. gage and up to 
and including 200-lb. gage; to close after blowing down at least 6 lb. on boilers carrying over 
200-lb. gage pressure. 

33. Each safety valve used on a boiler shall have a substantial lifting device, and shall have 
the spindle so attached to the disc that the valve disc can be lifted from its seat by means of the 
lifting gear a distance not less than one-eighth the nominal diameter of the valve seat. A safety 
valve used on hot-water heating boilers need not have lifting gear. 

34. Every safety valve shall be plainly marked, either by letters cast in the metal of the 
body or stamped on the body, with the words " Bevel Seat" or " Flat Seat *' according to the 
t3rpe of valve. Valves not marked with the words ** Bevel Seat" or ** Flat Seat," shall be deemed 
to have bevel seats. 

35. Every new pop safety valve shall be plainly stamped on the body witl^ figures showing 
the steam pressure at which it is set to blow. 

36. Every safety valve shall have the name or identifying trade-mark of the manufacturer 
plainly cast or stamped on the body. 

37. The seats and discs of safety valves shall be of non-ferrous material. A safety valve 
having either seat or disc of cast iron or steel shall not be used on a boiler for steam or water. 

38. Springs used in safety valves must not show a permanent set exceeding Vn '^- ten 
minutes after being released from a cold compression test closing the sining solid, coil to coil. 

39. The spring in a safety valve shall not be used for any pressure more than ten (10) per 
cent above or below the working pressure for which it was designed. 

40. Every safety valve used on a superheater, or discharging superheated steam, shall have 
a steel body with flanged inlet connection, and shall have the seat and disc of nickel composition 
or equivalent material, and shall have the spring exposed outside of the vajve casing so that the 
spring shall be protected from contact with the escaping steam. 

41. A safety valve used on a superheater shall be not larger than 3-in. size, and shall be 
connected near the outlet of the superheater. Two or more safety valves may be used on a super- 



BOILERS AND RULES FOR CONSTRUCTION 131 

heater, and one or more safety valves may be placed near the inlet of the superheater. The 
discharge capan^ty of the safety valves on a superheater shall not be included in determining 
the safety valves required for the boiler. 

42. During a hydrostatic test of a boiler, the safety valve or valves shall be removed or 
each valve disc shall be held to its seat by means of a light testing clamp and not by screwing 
down the compression screw upon the spring. 

43. A safety valve over 3-in. sice, used (or pressure greater than 15-lb. gage, shall have a 
flanged inlet connection. The dimensions of the flanges of safety valves shall conform to the 
American Society of Mechanical Engineers standard for the ocMresponding commercial pipe size. 

44. Fusible Plugs. Fusible plugs, if used, shall be filled with tin with a melting point be- 
tween 400 and 500 deg. Fahr. 

45. The least diameter of fusible metal shall be not less than }/i in., except for working 
pressures ci over 175 lb. or when it is necessary to place a fusible phig in a tube, in which cases 
the least diameter of fusible metal shall be not less than % in. 

46. Each boOer may have one or more fusible plugs, located as follows: 

^rintf In Horisontal Return Tubular Boilers — ^in the rear head, not less than 2 in. above the 
upper row of tubes, the measurement to be taken from the line of the upper surface of tubes to 
the center Of the plug and projecting through the sheet not less than 1 in. 

(b) In Horisontal Fhie Boilers — in the rear head, on a line with the highest part of the boiler 
exposed to the inroducts of combustion, and projecting through the sheet not less than 1 in. 

(c) In Locomotive Stationary Type or Star Water-tube Boilers — in the highest part of the 
crown dieet, and projecting through the sheet not less than 1 in. 

(d) In Vertical Fire-tube Boilers — ^in an outside tube, not less than one-third the length of 
the tube above the lower tube sheet, and projecting through the sheet not less than 1 in. 

(e) In Vertical Fire-tube Boilers, Corliss Type — ^in a tube, not less than one-third the length 
of the tube above the lower tube sheet. 

(/) In Vertical Submerged Tube Boilers — ^in the upper tube sheet. 

(g) In Water-tube Boilers, Horizontal Drums, Babcock A Wilcoxx Type — in the upper 
drum, not less than 6 in. above the bottom of the drum, over the first pass of the products of 
combustion, and projecting through the sheet not less than 1 in. 

{h) In Stirling Boilers, Standard Type — in the front side of the middle drum, not lees than 
4 in. above the bottom of the drum, and projecting through the sheet not less than 1 in. 

(i) In Stirling Boilers, Superheater Type — ^in the front drum, not less than 6 in. above the 
bottom of the drum, exposed to the products of combustion, and projecting through the sheet 
not less than 1 in. 

(j) In Water-tube Boilers, Heine Type — in the front course of the drum, not less than 
6 in. above the bottom of the drum, and projecting through the sheet not less than 1 in. 

(k) In Boib-Mumford Boilers, Standard Type — in the bottom of the steam and water 
drum, 24 in. from the center of the rear neck, and projecting through the sheet not less than 
1 in. 

(2) In Water-tube Boilers, Almy Type — in a tube or fitting exposed to the products of 
combustion. 

(m) In Vertical Boilers, Climax or HazeUon Type — in a tube or cent^ drum not less than 
one-half the height of the shell, measuring from the lowest circumferential seam. 

(n) In CahaU Vertical Water-tube Boilers — ^in the inner sheet of the top drum, not less than 
6 in. above the upper tube sheet, and projecting through the sheet not less than 1 in. 

(o) In Wickes Vertical Water-tube Boilers, in the shell of the top drum and not leas than 
6 in. above the upper tube sheet, and projecting through the sheet not less than 1 in.; located 
so as to be at the front of the boiler and exposed to the first pass of the products of 
combustion. 

(p) In Scotch Marine Type Boilers — in the combustion chamber top, and projecting through 
the sheet not less than 1 in. 



132 POWER PLANTS AND REFRIGERATION 



(^ 



In Dry Back ScoU^ Type Boilere — ^in the rear head, not less than 2 in. above the upper 
row of tubes, and projecting through the sheet not less than 1 in. 

(r) In Economic Type Boilers — in the rear head, above the upper row of tubes. 

(^ In Cast-Iron Sectional Heating BoUers — ^in a section over and in direct contact wiUi the 
products of combustion in the primary combustion chamb^. 

(0 In Water-tube Boilers, WorihingUm Tsrpe — in the front side of the steam and water 
drum, not less than 4 in. above the bottom of Uie drum, and projecting through the sheet not 
less than 1 in. 

(u) For other types and new designs, fusible plugs shall be placed at the lowest permissible 
water level, in the direct path of the products of combusticm, as near the primary oorobustian 
chamber as possible. 

47. Steam Gage. Ektch boiler shall have a steam gage connected to the steam q>aoe of the 
boiler by a siphon, or equivalent device, sufficiently large to keep the gage tube filled with water, 
and connected in such a manner that the steam gage cannot be shut off from the boiler ezoqjt 
by a cock with tee or lever handle, which shall be placed on the pipe near the steam gage. Con- 
nections to gages shall be made of brass, copper, or bronze compositicm pipe and fittings from the 
boiler to the gage. The handle of the cock shall be parallel to the pipe in which it is located 
when the cock is open. 

48. The dial of the steam gage shall be graduated to not less than 1}^ times the mftximinn 
pressure allowed on the boiler as stated in the certificate of inspection. 

49. Each boiler shall be provided with a ^^-in. pipe size connection for attaching the in- 
spector's test gage when the boiler is in service, so that the accuracy of the boiler steam gage 
can be ascertained. 

60. Water Glass and Gage Cocks. Each boiler canying over 15 lb. shall have at least one 
water glass, the lowest visible part of which shall be not less than 2 in. above the locations estab- 
lished for the top of fusible plugs in Par. 46, whether a fusible plug is used or not. Shutroff valves 
of the outside screw and yoke gate type are advised in both top and bottom connections to boiler 
to permit of blowing through either independently. 

(a) Each boiler shall have two or more gage cocks, located within the range of the visible 
length of water glass, when the maximum pressure allowed does not exceed 15 lb. per sq. in. 
except when such boiler has two water glasses, located not less than 3 ft. apart, on the same 
horizontal line. 

(&) Each b(Hler shall have three or more gage cocks, located within the range of the visible 
length of water glass, when the maximum pressure allowed exceeds 15 lb. per. sq. in., except when 
such boiler has two water glasses, located not less than 3 ft. apart, on the same horizontal line. 

51. Feed Pipe. Each boiler shall have a feed pipe fitted with a check valve, and also a 
stop valve or stop cock between the check valve and the boiler. Means must be provided for 
feeding a boiler with water against the maximum plressure allowed on the boiler. 

52. Stop Valve. Each steam outlet from a boil^ (except safety valve connections) shall 
be fitted with a stop valve. 

53. When a stop valve b so located that water can accumulate, ample drains shall be 
provided. 

54. Dandier Regulator. When a damper regulator is used, the boiler pressure pipe shall 
be fitted with a valve or cock,' and shall be connected to the steam space of the boiler. 

55. Lamphrey Fronts. Each boiler fitted with a Lamphrey Boiler Furnace Mouth Protector, 
or similar appendage, having valves on the pipes connecting them with the boiler, shall have 
these valves locked or sealed open, so that the locks or seals will require to be removed or broken 
to shut the valves. 

56. Bottom Blow-off. Each boiler shall have a blow-(^ pipe, fitted with a valve or cock, in 
direct connection with the lowest water space practicable. 

57. Valves on Return Pipes. The main retmn pipe to a heating boiler (gravity return 
system) shall have a check valve, and also a stop valve between the check valve and the boiler. 



BOILERS AND RULES FOR CONSTRUCTION 



133 



58. Wh^i there are two connected boilers (gravity return system), a check valve and a stop 
valve shall be installed in the branch pipe to each boiler, as shown in Fig. 35. 

59. Ptower Ratings for Classification. The horsepower of a boilef shall be ascertained upon 
the basia oi 3 hp. for each square foot of grate surface, if the boiler is used for heating purposes 



Stop Vdltfe 




Stop Valvo 



Fia, 35. Pbopeb Abbanqeicbmt or BLOwOnr Connbctionb to Two Boilers Opbbatbd on GB^vmr 

Return System. 



exclusively, and the safety valve is set to blow at 15 lb. or less. The horsepower of any boiler, 
whose safety valve is set to blow at over 15 lb., shall be ascertained on the basis of 2 hp. per 
square foot <^ grate surface, provided the grate surface is not in excess of 15 sq. ft.; if the grate 
surface is greater than 15 sq. ft., then the horsepower rating shall be based upon the heating 
surface, 10 sq. ft. of heating surface to be valued as equivalent to one horsepower. The engine 
power shall be computed upon a basis of a mean effective pressure of 40 lb. per sq. in. of piston 
for a simple engine; 50 lb. for a simple condensing engine; and 36 lb. for a compound condensing 
«igine, computed upon the area of the low pressure piston. The power rating of steam turbines 
shall be baaed on the builders' brake horsepower name plate rating when such information is 
available; €»* when not available and the steam turbine is direct connected to electric generat- 
ing apparatus the power rating shall be taken as the kilowatt rating of the generator times 1.34. 
In case other suitable means of determining the rating is lacking, the chief inspector of the State 
Boiler Department may cause such investigation to be made as is necessary to determine the 
normal capacity of the tmrbine, and his decision as to the rating to be allowed shall be finaL 

60. Annual Internal Inspections. The owner or user of a boiler which requires annual 
inspection, internally and externally, by the boiler inspection department or by an insurance 
company, as provided by Par. 1 of the Boiler Inspection Law, shall prepare the boiler for in- 
specti<Hi by cooling it down (blanking o£f connections to adjacent boilers, if necessary), removing 
aQ soot and ashes from tubes, heads, shell, furnace and combustion chamber; drawing c^ the 
water; removing the handhole and manhole plates; removing the grate bars from internally 
fired boilers; and removing the steam gage for testing. 

61. If a boiler has not been properly cooled down, or otherwise prepared for inspection, the 
boiler mspector shall decline to inspect it, and he shall not issue a certificate of inspection until 
^cient inspection has been made. 

62. In making the annual internal and external inspection under no steam pressure, as 
provided by Pars. 1 and 3 of the Boiler Inspection Law, the boiler inspector shall apply the 
hammer test to aU internal and external parts of a boiler that are accessible. 



134 POWER PLANTS AND REFRIGERATION 

63. All proper meamirements shall be taken by the boiler mspector, so that the maxiinuin 
working pressure allowed on a boiler will conform to the rules relating to allowable pressures 
established by the Boiler Code CommiUee of The American Society cj Mechaniail Engineers; 
such measurement to be taken and calculations made before a hydrostatic pressure test is applied 
to a boiler. 

64. The steam gage of a boiler shall be tested and its readings compared with an accurate 
test gage, and if, in the judgment of the boiler inspector, the gage is not reliable, he shall order 
it repaired or replaced. 

65. Annual External Inspections. The annual external inspection of a boiler, as provided 
by Par. 5 of the Boiler Inspection Law, should be made under allowable working pressure at or 
about six months after the annual internal inspection, except in the case of a boiler that is in 
service a portion of the year only, in which case the annual external inspection under steam 
pressure shall be made during such period of service. If a boiler or group of boilers is discon- 
tinued for an> reason other than defect, from service for long periods, a thorough internal and 
external inspectior under no pressure may be substituted for the regular external inspection 
under pressure. 

66. The boiler inspector shall attach an accurate test gage to a boiler to note the pressure 
which it shows, and compare it with that shown by the boiler gage, ordering the boiler gage 
repaired or replaced if necessary. 

67. The boiler inspector shall see that the water glass, gage cocks, water-column connections 
and water blow-offs are free and clear; also, that the safety valve raises freely from its seat. 

68. Fire doors, tube doors, and doors in settings shall be open to show as far as possible 
the fire surface, settings, tube ends, blow-off pipes and fusible plug. The boiler inspector shall 
note conditions and order changes or repairs if necessary. 

69. Hydrostatic Pressure Tests. When a boiler is tested with hydrostatic pressure the 
pressure applied shall not exceed one and one-half times the maximum allowable working pres- 
siu*e; except that a test pressure of 60 lb. shall be applied to the sections of boilers constructed 
entirely of cast iron (with the exception of their connecting nipples) and where the maximiun 
steam pressure does not exceed 15 lb. per sq. in. 

70. Cast-iron water boilers for heating buildings and water for domestic purposes, con* 
structed entirely of cast iron (with the exception of their connecting nipples) and subjected to 
a working pressure of less than 30 lb., shall be tested with a hydrostatic pressure of 60 lb. per 
sq. in. 

71. Cast-iron water boilers for heating buildings and water for domestic purposes, con- 
structed entirely of cast iron (with the exception of their connecting nipples) and subjected to 
a working pressure of 30 lb. or over, shall be tested with a hydrostatic pressure of twice the working 
pressure. 

72. Pipe boilers constructed entirely of pipe and pipe fittings may be tested to twice their 
working pressure. 

73. The boiler inspector, after applying a hydrostatic pressure test, shall thoroughly ex- 
amine every accessible part of the boiler both internally and externally. 

74. Efficiency of Joint The ratio which the strength of a unit length of a riveted joint 
has to the same imit of length of solid plate is known as the efficiency of the joint and shall be 
calculated as shown by the following examples: 

TJS. » tensile strength stamped on plate, in pounds per square inch. 
i » thickness of plate, in inches. 
b ^ thickness of butt strap, in inches, 
p = pitch of rivets, in inches, on row having greatest pitch. 
d = diameter of rivet after driving, in inches = diameter of rivet hole. 
a = crossHsectional area of rivet after driving, in square inches. 
8 = strength of rivet in single shear, as given in Par. 12 of these rules. 



BOILERS AND RULES FOR CONSTRUCTION 



135 



8 = Btr^igth cl rivet in double shear, as given in Par. 12 of these rules. 
c =s crushing strength of mild steel, as given in Par. 12 of these rules. 
n s number ol rivets in single shear in a unit of length of joint. 
iV s number of rivets in double shear in a unit of length of joint. 

75. EzAinpIe. Lap joint, longitudinal or circumferential, single-riveted; 
A » strength of solid plate » p X < X T.87 
B =» strength of plate between rivet holes = (p — d) < X T,S, 
C " shearing strength of one rivet in single shear '^ n Xs Xa, 
D « crushing strengUi of plate in front of one rivet ^ d X t X c. 




fa 




<i) ^ © 



Pio. Z6, Lap Joint, LoNorruDiNAL or Cibcumferential. Sinqle-Rivetkd. 



Divide B, C, or D (whichever is the least) by A, and the quotient will be the efficiency of a single- 
riveted lap joint as shown in Fig. 36. 



T.S. «- 56,000 lb. 

< — Ji m. = 0.25 in. 

p » l^in. « 1.625 in. 

d " H in. *= 0.6875 in. 

a » 0.3712 sq. in. 

a » 42,000 lb. 

12.890 (B) 

22,343 {A) 



e = 95,000 lb. 

A «= 1.625 X 0.25 X 65,000 « 22,343. 
B - (1.625 - 0.6875) 0.25 X 65,000 « 12.890. 
C « 1 X 42,000 X 0.3712 = 15,590 
D « 0.6875 X 0.25 X 96.000 «= 16,328. 

0.576 « efficiency of joint. 



76. BzJunple. Lap joint, longitudinal or circumferential, double-riveted. The vertical distance 
between th e two lines of rivet holes commonly termed the back pitch should be about 70 per cent of 
the pitch, ^e rivets being staggered. 

A « strength of solid plate = p X < X T.S, 

B «» strength of plate between rivet holes ■= (p — d) < X T.8. 

C « shearing strength of two rivets in single shear ^ n X 8 X a. 

D « crushing strength of plate in front of two rivets ^nXdXtXc, 

Divide B, C or D (whichever is the least) by A^ and the quotient will be the efficiency of a double- 
riveted Up jomt. (Fig. 39). 



TJ3. - 65.000 lb. 

I « A in- - 0.3125 in. 
p «2j^in. =2.875 in. 
d « Ji in. = 0.75 in. 
a » 0.4418 sq. in. 
a -* 42,000 lb. 

36,623 (B) 



49.414 (A) 



c « 95,000 lb. 
A « 2.876 X 0.3125 X 55,000 - 49,414. 
B = (2.875 - 0.76) 0.3125 X 66,000 = 36.523. 
C - 2 X 42,000 X 0.4418 « 37,111. 
2> « 2 X 0.75 X 0.3126 X 96.000 = 44,631. 

— 0.739 « efficiency of joint. 



136 



POWER PLANTS AND REFRIGERATION 



77. Example. Butt and double strap joint, double-riveted. 
A « strength of solid plate « p X < X T.5. 

B « strength of pl^te between rivet holes in the outer row — (p — rf) < X T.S. 
C » shearing strength of two rivets in double shear, plus the shearing strength of one rivet 

in single shear ^NXSXa+nXsXa. 
D » strength of plate between rivet holes in the second row, plus the shearing strength of one 

rivet in single shear in the outer row = (p — 2(i) < X T.S. -^ n X a X a. 
E ■■ strength of plate between rivet holes in the second row, plus the crushing strength of butt 

strap in front of one rivet in the outer row « (p — 2d) < X T.8, + d X 6 X c 




Fio. 37. Butt and Double Strap Joint, Doubls-Rivbtbd. 



F » crushing strength of plate in front of two rivets, plus the crushing strength of butt strap 

in front of one rivet '^ N Xd Xt Xc -{■ n Xd Xh Xc, 
O ■■ crushing strength of plate in front of two rivets, plus the shearing strength of one rivet in 
single shear '^NXdXtXc-^-nXBXa. 
Divide B, C, 2>, E, F or O (whichever is the least) by A, and the quotient will be the efficiency 
of a butt and double strap joint, double-riveted, as shown in Fig. 37. 



.5. - 65,000 lb. 
< « J^ in. = 0.375 m. 
6 - A in. = 0.3125 in. 
p - 4 J^ in. « 4.875 in. 
d « >g in. = 0.876 in. 



a = 0.6013 sq. in. 
a « 42.000 lb. 
S = 78,000 lb. 
c = 95,000 lb. 



Number of rivets in single shear in a unit of length of joint ■■ 1. 
Number of rivets in double shear in a unit of length of joint » 2. 

A « 4.876 X 0.375 X 66,000 = 100,547. 

B « (4.876 - 0.875) 0.375 X 55,000 = 82,500. 

C = 2 X 78,000 X 0.6013 + 1 X 42,000 X 0.6013 = 119,057. 

D « (4.875 - 2 X 0.875) 0.376 X 55,000 + 1 X 42,000 X 0.6013 - 89,708. 

E - (4.875 - 2 X 0.875) 0.376 X 55,000 + 0.876 X 0.3125 X 95,000 =• 90,429. 

F « 2 X 0.875 X 0.376 X 95,000 + 0.876 X 0.3125 X 95,000 = 88,320. 

- 2 X 0.875 X 0.375 X 95,000 + 1 X 42,000 X 0.6013 = 87,699. 



82,500 (B ) 
100,547(ii) 



0.820 s= efficiency of joint. 



78. Example. Butt and double strap joint, triple-riveted. 
A = strength of solid plate « p X < X T.S. 

B — strength of plate between rivet holes in the outer row = (p — d) < X T.S. 
C " shearing strength of four rivets in double shear, plus the shearing strength of one rivet in 
sin^e shear »^X5Xa-f-nX«Xo. 



BOILERS AND RULES FOR CONSTRUCTION 



Fig. 38. Bott ani 

" Btrensth of plate between ri 
rivet in single shear in tl 

" etrenKth of plate between ri 
strap in front of one ri 



I DouBLS Strap Joint, Tbipis-Rivktbd. 



Bt holes in the second row, plus the ■} 
1 outer row - (p ~ 2d) t X T.3. + n X » X a. 
:t holes in the second row. plus the cruahinB ntreoeth of butt 
in the out*r row - (p - 2d) X T.S. +d Xb Xc 
•• eruahine atrenatli of plate in front of tour rlvota, plus the cmahing Bticnsth of butt strap 
in front of one rivet - N Xd Xt Xc + n Xd Xb X e. 
G ~ CTUshinc strensdi of plate in front of four rivets, plus the shearing strength of ope rivet 
in single shear -NxdXtXe+nXaXa. 
Divide B, C. D, B.Ftaa (whichever is the least) by A, and the quotient will be the el 
a batt and double strap joint, triple-riveted, bs shown in Fig. 38. 



T.S. - 56,000 lb. 

I - !^ in. - 0.376 in. 
b - A i°- - 0.3125 in. 
p - 6>i in. - 6.5 in. 
d - H in- - 0-8125 in. 



a - 0.6186 sq. it 

t - 42,000 lb. 
S - 78,000 lb. 



SISSLC SlVETEO LAP JOIHT, 

Bfnasucfss pes cesr. 




emit AT'i^-y 

DOUBLE RIVFTEa UP JOIUT, 
EmCIEMCr 72 PER CENT. 



TmPLS PlfETeO BUTT-STRAP JOIHT, 




- acCTTOH AT'K- 



aSCTlOH AT'X- 
^ADBUPLB RUTTEO BUTT-STRAP JOIST. 
DOUBLE COVERS 
EFFICIEflCr H PER CEUT 



Fio. 39. TmcAL LoirarrDDiMAi. Ritbtxd Boilbb Joiotb. 



140 



POWER PLANTS AND REFRIGERATION 



Number of rivets in single shear in a unit of length of joint >■ 1. 
Number of rivets in double shear in a unit of length of joint * 4. 

A - 6.5 X 0.375 X 55.000 - 134.002. 

B - (6.5 - 0.8125) 0.375 X 55.000 - 117,304. 

C - 4 X 78,000 X 0.5185 + 1 X 42,000 X 0.5185 - 183,549. 

D = (6.5 - 2 X 0.8125) 0.375 X 55,000 + 1 X 42.000 X 0.5185 - 122.323. 

E - (6.5 - 2 X 0.8125) 0.375 X 55,000 + 0.8125 X 0.3125 X 95,000 - 124,667. 

F - 4 X 0.8125 X 0.375 X 95,000 + 1 X 0.8125 X 0.3125 X 95,000 - 139.902 

G - 4 X 0.8125 X 0.375 X 95,000 + 1 X 42,000 X 0.5185 - 137,558. 



117,304 (B ) 
134,062 {A) 



0.875 >• efficiency of joint. 



TABLE 25 

LAP JOINTS 



Thiek. 

of 
Plate, 
Indies 



Diem. 

of 
Rivet, 
Indies 



Center 
of Hole 
to Edfe 
of Plate, 

Inches 



SiNGLS RiVBTBD 



Pitdi, 
Inches 



Lap, 
In»es 






m 


Wa 


2K 




m 


li! 


I 


i|i 


2 


ii| 


2H 


211 


\\ 


IH 


2K 


8 


1 




^ 


8A 


1 


»A 



Efficiency 
of Plate, 
Per Cent 



67.1 
66.6 
66.2 
66.8 
66.6 
66.4 
64.0 



DOUBLB RiVBTBD 



Diam. 

of Rivet, 

Indies 



S 



1: 



I! 



Pitch. 
Indies 



Lap, 
Inches 



Rows, 
Inches 



IS 

2H 



KinciBucy 
of PUte, 
Per Cent 



72.7 
72.3 
72.0 
71.1 
70.8 
71.4 
70.6 



TABLE 26 

TRIPLE RIVETED BUTT STRAP JOINT 



Thick- 
ness of 
Plate. 
Indies 



EHam. 

of 
Rivet, 
Indies 









Pitch of 
Rivets 

in 
Inches 



8H 

3H 
3H 
3H 
SH 

8H 



6H 
6H 
6H 

7 



of 



Width 
Outside 
BuU 
Strap 



10 K 

10 H 



Width 
of Inside 
Butt 
Strap 



14 
14 

14 K 
14^ 

16K 

16^ 

16H 



Thick- 
ness of 
Cover 
Straps 



M 
ft 

I 



Vertioa 

or 

Trans. 

Pitch 



Edfe of 
Butt Strap 
to Center 
of RiveU 



Efficiency 

of 

Plate 

Per Cent 



88 

88.6 

87.9 

88.4 

87.9 

87.6 

87.8 



TABLE 27 

QUADRUPLE RIVETED BUTT STRAP JOINT 



y% 



IH X 14! 
7^x145 
7^x15 
8 zl6 
8 xl6 



10 H 
lOH 



20 ^ 

20 H 
22 

22H 
22H 



94.4 
94.5 
94.2 
94.1 
94.1 



BOILERS AND RULES FOR CX)NSTRUCTION 



141 



TABLE 28 

LAP-WELDED STEEL OR CHARCOAL IRON BOILER TUBES 

Table of Standard DimengioDa 



DlAMBTEB 


Nom. 
rhiftk- 


Nflwwi 

BWIn 

Gage 

No. 


CmCUlfFERBNCB 


TBAHirirMljf ARBA8 


Length of 

TUBB PER 

Sq.Pt. OF 


Nom. 


External, 
IndMa 


Intarnal. 
In. 


External, 
Inches 


Internal, 
Inches 


External, 
Sq. In. 


Internal, 
Sq. In. 


Metal. 
Sq.In. 


Ex- 
ternal 
Burfaoe, 
Feet 


In- 
temal 
Surfaoe, 
Feet 


Wgt. 

perFL 

Lb. 


1 


0.810 
1.060 
1.810 
1.660 
1.810 
2.060 
2.282 
2.682 
2.782 
3.010 
8.260 
3.610 
8.782 
4.282 
4.704 


9.096 
.096 
.096 
.096 
.096 
.095 
.109 
.109 
.109 
.120 
.120 
.120 
.134 
.184 
.148 


13 
18 
18 
13 
13 
13 
12 
12 
12 
11 
11 
11 
10 
10 
9 


8.142 

3.927 

4.712 

6.498 

6.288 

7.069 

7.864 

8.639 

9.426 

10.210 

10.996 

11 .781 

12.666 

14.187 

16.708 


2.645 

8.830 

4.116 

4.901 

5.686 

6.472 

7.169 

7.964 

8.740 

9.466 

10.242 

11.027 

11.724 

18.296 

14.778 


0.786 

1.227 

1.767 

2.406 

8.142 

3.976 

4.909 

6.940 

7.069 

8.296 

9.621 

11.046 

12.666 

16.904 

19.636 


0.515 

.882 

1.847 

1.911 

2.673 

8.833 

4.090 

6.086 

6.079 

7.116 

8.847 

9.676 

10.989 

14.066 

17.879 


0.270 

.844 

.419 

.494 

.669 

.643 

.819 

.906 

.990 

1.180 

1.274 

1.869 

1.627 

1.838 

•2.266 


8.819 
8.066 
2.647 
2.188 
1.909 
1.698 
1.628 
1.889 
1.278 
1.176 
1.091 
1.018 
0.966 
.849 
.764 


4.716 
3.608 
2.916 
2.448 
2.110 
1.864 
1.674 
1.609 
1.378 
1.269 
1.172 
1.088 
1.024 
0.902 
.812 


0.90 




1.16 


J 1^ 


1.40 


1 W 


1.66 


2 


1.91 


2K 


2.16 


aS .. 


2.76 


2H :::::::::: 


8.04 


a 


3.83 


8W 


8.96 


svS 


4.28 


8H 


4.60 


4 


6.47 


4H 


6.17 


6....::::::.:...:.. 


7.68 







CHAPTER V 

MECHANICAL STOKERS 

In all types of mechanical stokers the coal is fed to the fire automatically from a hopp^ 
placed, with practically all tjrpes of stokers, in front of the boiler. Unless, however, the coal is 
automatically fed to the stoker hopper from overhead storage and the ash removed automatically 
there is no considerable saving in labor with a mechanical stoker installation. Examples of 
modem stoker installations will be found in the Chapter on '' Arrangement of Steam Power Plants." 

The mechanical stokers permit the use of cheap fuels with good economy and when properly 
installed give practically smokeless combustion and are frequently installed for this reason, 
regardless of other considerations. 

The smokeless combustion feature is due to the more even and continuous firing than is 
obtained with the intermittent firing of hand-fired furnaces. 

There is a tendency with all types of stokers to cau^e the loss of some fuel by sifting throu^ 
the grate into the ash pit. Suitable arrangements, however, may be made to recover or reclaim 
this fuel. 

The amount of sif tings depends upon the type of stoker and the degree of fineness of the coaL 
With bituminous slack, the amount sifting through a chain-grate stoker frequently amounts to 
10 per cent or over, and would represent a considerable loss unless recovered. 

If run of mine bituminous coal is to be used, a coal crusher is an essential part of the equip- 
ment when stokers are to be used. 

" Many of the successful stokers of to-day are utilizing a forced draft for their operation. 
At one time it was assumed that this class of stokers did away with the necessity of a stack, except 
for carrying off the gases of combustion. As combustion rates increased, however, and it was 
necessary to supply more and more blast, it became evident that, from the standpoint of pro- 
tection to furnace brickwork, a draft suction was necessary as well as a blast. In view of the 
enormous heat now developed in stoker-fired furnaces and the great weight of the gas passing 
over the boiler-heating surfaces, it is now generally accepted that some means must be provided, 
whether by natural or induced draft, to remove these gases from the furnace promptly in order 
to prevent a 'soaking up' action of the heat by the furnace brickwork. To assure such 
removal, the means provided should be such as to give a draft suction throughout all parts of 
the setting under any conditions of service." 

With hand-fired boilers, one fireman can take care of the coal and ashes for approximately 
300 to 400 boiler horsepower. This assumes that the distance from the coal pile to boilers is 
not over 100 feet and that he fires direct from the truck or barrow. 

If the arrangement is such that the coal must be handled more than once, a fireman and 
helper will be required for each battery of 500 boiler horsepower. With a complete equipment 
of automatic coal and ash-handling machinery in conjunction with automatic stokers, one fire- 
man can take care of four batteries of 500 boiler horsepower units. 

The following data were compiled by R. T. Hale from the reports made by four hundred 
members of the Steam Users* Association of New England: 

" Under average conditions, one man in addition to a night man can run an engine and fire 
up to ten tons of coal per week. 

" For thirty tons it requires an engineer, one day and one night man." The capacity of 
such a plant would be approximately 200 horsepower. 

142 



MECHANICAL STOKERS 143 

" For a 600 hOTsepower plant it would require an engineer, two day men and one night 

From aD inveetigation of 600 small plants the average cost of handling coal was 48 cents 
per ton, with a nuudmum of 71 cents and a minimum of 26 cents per ton. 

When coal was moved by wheelbarrow the coet was 1.6 cents per ton pet yard for the first 
five yards and about 0.1 cent per ton for each additional yard. 

CUssiflcatioti ol Stokera. Mechanical stokers are of three general types: fH overfeed, 
(2) undnrfeed, {3} traveling or chain grate. 

Overfeed Stokers in general may be divided into two classes, the distinction being in the 
directicKi in which the coal is fed relative to the furnaces. In one cIshb the coal is fed into hoppers 
at the boot end of the furnace on to grateswith an inclination downward toward the rear of 



Thb Ronet Overfeed Btokeb. 



pio. 2. The Joins UNVBarEBD Stozeb. 

about 46 d^reea. These grates are reciprocated, being made to take alternately level and 
inchned positions, and this motion gradually carries the fuel as it is burned toward the rear and 
bottom of the furnace. At the bottom of the grates flat dumping sections are supplied for com- 
pleting the combustion and tor cleaning. The fuel is partly bumed or coked on the upper portion 
of the grates, the volatile gases driven off in this process for a perfect action being ignited and 
bumed in their passage over the bed of bummg carbon lower on the grates or on becommg 
mixed with the hot gases in Ow furnace chamber. In the second class the fuel is fed from the 



144 POWER PLANTS AND REFRIGERATION 

sidee of the furnaoe for its full depth from front to rear on to grates inclined toward the center 
of the furnace. It is moved by rocking bars and is gradually carried to the bottom and center 
of the furnace as combustion advances. Here some type of a so-called clinker-breaker removes 
the refuse. 

The Btmey, WiOemton, Acme, Murphy and DetroU are examples of this ohws. 



Flo. 3. OOMFLETB IHSTAUAITON OP m JOHIB StoKIB SBOWINa FAN. AlH DUITTB ANS BTMAK OFBKATnO 

Underfeed Stokers are either horiEontal or inclined. The fuel ia fed from underneath, either 
continuously by a screw or intermittently by plungers. The principle upon which these stokers 
base their claims for efficiency and smokeleauiess is that the green fuel is fed under the coked 
and burning cosl, the volatile gasea from this fresh fuel being heated and ignited in their psnage 
through the hottest portion of the fire on the top. In the horiiontal clasBea of underfeed stokera, 
the action of a screw corriea the fuel back through a retort, from which it paasea upward as the 
fuel above is consumed, the oah being finally deposited on dead plates on either aide of the retort, 
from which it can be removed. In the inclined class, the refuse is carried downward to the rear 
of the furnace where there are dumping platce, as in some of the overfeed types. 

Underfeed stokers are ordinarily operated with a forced blast, this in some cases being 
operated by the same mechaniain as the stoker drive, thus automatically meeting the require- 
ments (rf various combustion rates. 



MECHANICAL STOKERS 




Fia. 4. Tas T&TLOB Stokib. 



146 POWER PLANTS AND REFRIGERATION 

The Jtmea, Ameriam, Combuition Engineering Co. Type E, Tayltr and Watii^houte ara ex- 
am plea of this class. 

TiATeling at Chain Grates are of the class best illustrated by chain-grate stokerH. As implied 
by the name, these consist of endless grates composed of short sections of bars, pasBing over eprock- 
eta at the front and rear of the furnace. Coal is fed by gravity on to the forward end of the 



Fio. 6. The Orbbn CHAiH-OnATB Stoker. 

grates through suitable ho)ii>ers, is ignited under ignitioD arches, and is carried with the grate 
toward the rear of the furnace as its combustion progresBes. When operated properly, the 
combustion is completed as the fire reaches the end of the grate and the refuse is carried over 
this rear end by the grate in making the turn over the rear sprocket. In some caaee auxiliaiy 
dumping grates at the rear of the chain grates are used with success. 

Chain-grate stokers in general produce less smoke than either overfeed or underfeed types, 
due to the fact that the^e are no cleaning periods necessary. Such periods occur with the latter 
types of stokers at intervals, depending upon the character of the fuel used and the rate com- 
bustion. With chain-grate stokers the cleaning is continuous and automatic, tmd no periods 
occur when smoke will necessarily be produced. 

In the earlier forms, chain grates had an objectionable feature in that the admission of large 
amounts of excess air at the rear of the furnace through the grates was possible. This objection 
has been largely overcome in recent models by the use of some such device as the bridge-wall 
water-bon and suitable dampers. A distinct advantage iA chain grates over other typee is thai 



MECHANICAL STOKERS 147 

they can be withdrawn from the furnace for inspection or repairs without interfering in any way 
with the boiler setting. 

This class of stoker is particularly successful in burning low grades of coal running high in 
ash and volatile matter which can only be burned with difficulty on the other types. 

Tlie B, & W; Green and DuhUh are examples of this class of stoker. 

Cost of Mechanical Stokers. The following costs are based on bids received during 1916 for 
four stokers to serve four 500-horsepower water-tube boilers, natural draft: 

Cost per rated boiler horsepower $2.90 to $3.25 



CHAPTER VI 

SDPBRHBATERS AKD ECOHOUIZERS 

SUPERHEATERS 

Experience has demonstrated, particuUrly with steam turtiines, that a moderate degree of 
superheat (from 100 to 200 degs. F.) leads to ecoDomy aad this has become standard practice in 
ateam-turbine-drivea plonta. The per cent reduction in the water rate trf tutbinea due to Tarious 
degrees of superheat ia given in the Gh^>ter on " Steam Tuifoines." 







With a properly designed superheater, the increaae in the fuel consumption 
produce a given weight of steam ia approximately as follows: 





,H. 


1 


S.l 



The superheating tube surface is cwlinarily placed within the boiler setting in such a way 
that the products of combustion for generating saturated steam are also utiliced for superheatr 
i)^ the steam. See Pigs. 24, 25, 26 and 29 in the Chapter on " B<nleis," etc. 

Amount of Superheating Surface. The amount of superheating surface necessary to give 
the degree of superheat desired depends primarily on the temperature of the gas in contact with 
the superheating surface, the conductivity of the tubes and the velocity of the steam and gases 
through and over the surface. The average temperature of the gases in contact with the supo^ 
heater tubes will depend oo the location. The following data may be used for proportioning 
148 



SUPERHEATERS AND ECONOMIZERS 149 

superiMAttog aurface located inside the hmiet setting, for superbeata, 100 to ISO degs. and 135- 
Ib. gage, uaaog mild steel tubes; 



iifcniiM*.. 

sthmmtmar 



The iDstallation of a euperheator causes an addittonal draft lose of ^qnoumatdy O.IS ia. 
wata when boiler is opoated at ISO per cent rating. 



Fia. 2. tarn Hbinc anpsKHHATia. 

Vet SaTfaig Due to SuperheatlnK. Assume the following data: Initial steam pressure, 
150 lb. absolute; final temperature of feed water, 200° F.; water rate of turbine witii saturated 
■team, 16 lb. per kw.-bour. 

The heat required to generate 1 lb. of steam tor this condition is 86*,9 + (358.5 - 200) - 
1023.4 B.t.u. 

Assume that the steam is superiieated 120° and that the water rate of the turbine wiU be 
reduoed 1 per cent for each 12 de^. kA superheat. 

The additional heat required per lb. of steam is 0.65 X 120 - 77 B.t.u. The new water 
rate of the turbine is 16 — (0.10 K 16) = 14.4 lb. per kn.-hour. 

Using saturated steam, the heat input required is 1023.4 X 16 - 16,374,4 B.t.u. per kw.-hour. 

Using superheated steam, the heat input required is 1100.4 X 14.4 « 15,845.8 B.t.u. per 
kw.-hour. This gives a difference of 629 B.t.u. in favor of superheating or a saving of fuel of 

529 
- ■ ■ ■ - - 0.032 or 3.2 per cent. 

ECONOMIZERS 
An econoroiser is a device tor heating the feed water by means of the waste flue gases. They 
ore usually made up of vertical rows of 4-in. cast-iron tubes, approximately ft. in length, attached 
to cross headers at top and bottom, the whole being enclosed by either a sheet-iron or brick 
CMing (Fig. 3). 



150 



POWER PLANTS AND REFRIGERATION 



The feed water ent^s the econtmiiser at the end farthest from the bwler and flowa in the 
opposite direction to the gases (counter flow). Each tube la provided with a scraper, which is 
operated fiotn the outside to remove the accumulatioD of soot from the tubes in order that the 
beat-transmisBiiin efGciency may not be impaired. A by-pass around the economisere should 
always be provided in cn^er that they may be cut out 4^ service for repsirs without interfering 
with the operation of the boilera. 




Fia. 3. Qbsbk F'lrm. Eoonduebk. 
LongltudlnBl Elevation and 9ectloD oT Grrea'i EoonomiEsr^ (1) 9ecUon la Plane or Top AOTOa Branrii 
Pipe: (2| Section througb Pipes and Hand-Hoht 1Mb; (3) Section between Pipes and (4) FroDt Etevatloa 
with Sectional Covering In Place. 

The saving effected by the installatioci of an economiier will depend upon the initial tem- 
perature of the teed water entering the economizer and the temperature of the flue gases. The 
higher the rate of evapotstion, the higher will be the temperature of the flue gases and attendant 
kies, consequently the saving will be greatest with boilers running overloaded. 

The percentage of saving may be calculated in the same manner as given for feed-water 
beaters. (See Table 1, Chapter on " Feed -Water Heatere.") The percentage of saving duo 
to the economizer will be less in plants which have feed-water heaters. 

The temperatures of the gases in the furnace an3 the several passes over the tubes ^own in 
Fig. 4 were t^en from a test on a water-tube boiler containing 6000 square feet of surface arranged 
in U rows of 18-ft. tubes, each 4 in. in diameter. The temperatures were obtained by means of 



SUPERHEATERS AND ECX)NOMIZERS 



151 



an electrical pyrometer when operating the boiler at about nominal load, and show that the first 
five rows in the first pass absorbed 55 per cent of all the heat recovered, and that the second 
and third passes abstracted less than 16 per cent of the heat absorbed by the boiler. 

Tests made on another boiler gave the relations almost identical as shown in Fig. 5. 

Obviously, it does not pay to r^uce the temperature of the flue gases below 600 or 700 
d^rees by means of the boiler surface alone. But, on the other hand, to throw the gases away 
at this tenaperature involves a loss of 25 to 40 per cent of the total heating value of the fuel. 

The remedy is to substitute, for the additional boiler smface, economizer surface, which, 



000 


"^ 






— "■ 








—"" 


















"■"" 






^^" 










































400 


[ 








































900 


\ 
































— 








too 


\ 






































rf 100 


i\ 








































m MOO 




V 






































O 000 




v. 






































^ 800 




^ 






































S 700 

9 AAA 






s. 




































9 000 
§ BOO 

C 400 
% 300 






X 


















— 













\ - 












V 
















■ — 

























































































9. sOO 
1 100 
^1000 

9 dWMk 












^ 










































w. 






































^ 


























• 000 

1 aoo 

S TOO 
^ 000 




















»^^ 


o 




















































































— c 









^^^ 


— 






MO 


























1 






— ^ 








fOO 


































^ " 


— > 


900 
800 

100 

< 































































<— ^ 


— 1 — 


=Ti 


nt 


^»w 






--*j<t -S«Ton 


y Piww- 


^ ThjnJ I'ttsa 


I 


1 





9 



P 


9 

we 



lent 


i 

ofB. 



oOei 


6 
Tot 



leSn 



ufao 





7 
Med 



Ot< 


90 
•r. 


M 


100 



FiQ. 4. CuBVB OF Temperatorb Drop Through Passes of Boiler. 

because o( the lower temperature of its contents, and consequently greater temperature differ- 
ence between the two sides of the heating surface, absorbs heat more rapidly than does the sur- 
face in the last pass of the boiler. Moreover, the fixed charges upon the economizer surface are 
less, square foot for square foot, and it is therefore able to show much higher returns upon the 
investment than would additional boiler surface. 

Increase in Temperature of Feed Water Due to the Use of Economizer. 
Let (i = initial temperature of feed water degs. F. entering economizer, 

(s = final temperature of feed water degs. F. leaving economizer. 
Ti = initial temperature of flue gas degs. F. 
Tt — final temperature of flue gas degs. F. 
X = rise in temperature of feed water (tt — (i). 
D « algebraic mean temperature difference between feed water and flue gas 

{Ti - tt) -h {Tt - (i) 



w = pounds of feed water per boiler horsepower-hour. (Approx. 30.) 

W = weight of flue gases per boiler horsepower-hour, pounds. (Approx. coal per 

b.hp.-hour X 20.) 

c = specific heat of flue gas. (Approx. 0.20.) 

wx - B.t.u. absorbed by feed water per b.hp.-hour. 

eW {Ti — Tt) = B.t.u. given up by flue gas per bJip.-hour. 

wx 

Then Ti - T, - --, 

cW 

and D - Ti - ^ - X ^ Zr (^) 

2cW 



SUPERHEATERS AND ECONOMIZERS 



153 



S « square feet of economizer surface per boiler-horsepower. 

= 3.5 to 5. 
U s unit heat transmission of economizer tubes or B.t.u. transmitted per sq. 
ft. per deg. difference in mean temperature per hour. 
UD S s B.t.u. transmitted per hour per b.hp. 



xto 
UD S ^ zw ,\ S ^ — 

UD 

Combining equations (1) and (2) 



(2) 



S (Ti - h) 






(3) 



Substitutiiig w - 30, e -i 0.20 and U " 3.3 



s (r, - <,) 



(4) 



•■-(^^) 



s 



This is the empirical formula proposed by the Green Fud Economizer Co, 

It is not advisable to reduce the temperature of the flue gas below 250** F. as the resulting 
condensation of the water vapor on the tubes, together with the sulphurous gases, produces rapid 
corrosion. 

As usually proportioned, the temperature of gas leaving the economizer is from 250^ to 325° F. 






^ 



24 



22 



20 



^ § c 



18 



16 

o 
^/4 



i5. 



^. 



10 



8 













\ 












i 












• 




^ 




Vat 












11 




iOo/7^N 










i 


<l 1 ^ 






^**\ 




— 


' 1 




rA 


'.^'^ 






\ 




^ 










^ 




X 


/ « 


) 








\ 






i 


ZCONOMIZER PERFORMANCE 
DURING PEAK LOAD 


^ 


J 


.^ 













1700 



WOO 



1500 



1400^ 
1300 t 



t200 



5 



HOO 5 
lOOO^ 



900 
300 



1:30 2:30 3:30 4:30 



530 

Time. 



6:30 7:30 8:30 9:30 



700 



FiO. 6, 



154 



POWER PLANTS AND REFRIGERATION 



The average heat transmission of economizer tubes (U) in B.t.u. per sq. ft. per hour per 
degree difference in average temperature between the gases and water is given by C, S, Dow 
as foUowa: 

TABLE 1 

HEAT TRANSMISSION OF ECONOMIZER TUBES 



Initial Temperature of Flue Gas 


Unit 
Tranamiarion 

(to 


300 


2.25 


400 


2.75 


500 


8.00 


600 


8.25 







The average value of C/ is increased as the velocity of the gas over the tubes and the velocity 
of water through the tubes increase, which is in accord with the results obtained with hot-blast 
heaters (Volume I). 

Fig. 6 represents the performance of a large electric station during the period of the after- 
noon peak load and shows the increase in the value of {/ as the load on the plant is increased. 

The data given in the following table are taken from the practice of one economizer company: 



TABLE 2 



Temperature of Entering Flue Gas 



450*»to 600°. 
600°to700'». 



Temperature 
Rise of 




Feed Water 


Square Feet of 


per 100 
Degrees 


Heating Surface 
per B.hp. 


Reduetion 




in Temp. 




of Gaa 




eo** 


4 


66*» 


4Hto5 



TABLE 3 

RESULTS FROM TESTS OF GREEN ECONOMIZERS 
(Compiled from Tests Green Fuel Economizer Co.* 9 Catalog) 



Number of Tests 


Rate of 
Combus- 
tion, Lb. 
Dry Coal 
perSq.Ft 
Grate per 
Hour 


Equiva^ 

lent Evap. 

from and 

at 2 12'' per 

TJ>.Dry 
Coal 


Ratio 

Heating 

Surface 

to Grate 

Area 


Temperature 
Gasbs 


Temperature 
Water 


Temp. 
Rise 


Enfg 
Econ. 


Leav*g 
Econ. 


Ent'g 
Econ. 


TiPaving 
Econ. 


Feed 
Water 


1 


15.2 

12.57 

22.4 

■ • • • • 

21.86 


11.2 
11.59 
- 9.59 

• • • • • 

6.79 


62.5 
49.5 
19.5 

• • • • 

• • « • 

84.2 


435 
416 
620 
548 
603 
537 


279 
254 
293 
296 
325 
826 


84 

40 

101 

96 

93.5 
71.2 


196 

165.4 

237 

200 

208.8 

203.4 


112 


2 


121.4 


3 


136 


5o 


104 


56 


110.8 


6 


182.2 







SUPERHEATERS AND ECX)NOMIZERS 



156 



TABLE 4 

RESULTS FROM TESTS OP STURTBVANT ECONOMIZERS 
(Compiled from B. F, SturUpani Co,'b Catalog) 



• 


TmPUUTURB, Gasbb 


Tbmpbratubb, Watbh 


Temperature 


Number of Test 


Entering 
Eoonomiser 


Leaving 
Economizer 


Entering 
Economize 


Leaving 
Economizier 


Rise, Feed 
Water 


1 


660 
676 
470 
600 
460 
440 
626 


276 
290 
280 
240 
200 
220 
226 


180 
160 
180 
110 
90 
120 
180 


840 
820 
260 
230 
230 
236 
820 


160 


2 


160 


8 


180 


4 


120 


6 


140 


6 


116 


7 


140 







Example. Determine the final temperature of the feed water in a steam-turbine-driven plant 
of 1000-kw. capacity. 

Water rate of main unit 20 lb. per kw.-hour. Steam-driven auxiliaries use 10 per cent of the steam 
required for main unit. Jet type condenser used, vacuum maintained 28'' hg., corresponding tem- 
perature 101^ F. Open type heater using exhaust steam from the auxiliaries and feed water drawn 
from the condenser hot well. 

Feed water leaving open heater to be passed through an economizer. 

Assume a ** terminal difference*' for the final temperature of condensing water of 10^ F. Tem- 
perature of hot well and initial temperature of feed water, 101 — 10 «= 91® F. 

From the data given in the Chapter on "Feed-Water Heaters," it is readily calculated that for an 
assumed radiation of 10 per cent in the open heater that the temperature of the feed water leaving the 
heater will be about 199® F. This is the initial temperature of the feed water entering the economizer. 

Assuming a combined efficiency of 0.60 for the boiler plant and a calorific value of 13,000 B.t.u. 
per lb. for the fuel, 4.3 lb. of coal are required per b.hp.-hour. If 20 lb. of air are used per lb. of coal, 
the wdght of flue gases per b.hp. per hour is: TT «= 20 X 4.3 « 86 lb. 

Assume, w » 30, c = 0.20, S « 3.6, U = 3.3, Ti - 660® F., and h « 199. 

Substituting the above values in (4), 

3.5 (560 - 199) ^„ , . . , , J . . 

= 88 degs. nse m temperature of feed water m economiser. 



9.1 H- 



m^h'^ 



.*. fi « li + ^ ■■ 287® F. temperature of water entering the boilers. 

The final temperature of the flue gas leaving the economizer may be approximated by first solving 
for D in equation 2: * 

zw 88 X30 



D - 



US 3.3 X3.5 



228. 



228 



650 - 287 4- 3^2 - 199 



, from which Tt «= 392® F. 



Loss of Draft Thiotigh Economizers. Tests No. 5a and 56, Table 3, were made on the Man- 
hattan Power Station, IrUerborougk Rapid Transit Company^ New York City. 

The economizer in question contained 512 tubes 10' 0" long, 4-9/16" outside diameter, 
economizer heating surface per rated b.hp., 3.25 sq. ft. 'The clear area through economizer 
being nearly 3 sq. in. per b.hp., which is somewhat greater than the standard practice. The 
flue area is sH^tly greater than this. 

The loss of draft through economizer for test 5a, when' the boilers were being operated at 
6 per cent below rating, was 0.16" water, and 0.23" water for test 56, when the boilers were 13 
per cent overloaded. 



X ^^^^x\\\\\^^v\\^^«E=:^ 



I* v^r.. .•-."!»■ 




c 

N 

'i 

o 

c 
o 

CI 



c 

o 

i 

■o 

o. 
a. 

*3 

O* 
V 

o 
U 

J ^" 

H 

'o. 

Q. 

•> 



r 

o 

"a 
O 



c 

C9 



o 



158 



POWER PLANTS AND REFWGERATION 



The following results were obtained at the Ladede Power CoJs Plant, St. Louis. The height 
of stack refers to the height above the economizers. 

TABLE 5 



Base of ISO-foot sUdc 

Breeehinf between boiler and eecmomiier 

End oi intermecUate baflle 

Combuaticm rhamhw (eetimated) 

Purnaee (estimated) 



Temperature 




FlueGaaea, 


Draft. 


Defxees 


Inebca 


Fahrenheit 




800 


1.25 


486 


0.75 


770 


0.60 


1.600 


0.87 


2.100 


0.25 



TABLE 6 

GENERAL DIMENSIONS OF ECONOMIZERS 

Height over gearing 13' 6K". Height over aeetion 10" 2K". Outside diameter of tubes 41'. hMting surfaee 

tube 12 square feet. 



Number Tubes Wide 



6. 

8. 
10. 
12. 



8 

8 

12 

16 

20 



4' 10" 
4 10 
7 8 
9 8 
12 1 



DlMKNSIONS 
INSIDB 

Waub 



•i 



n 






8' 4" 
4 8 

6 

7 4 
3 8 



i 



CO 



4' 1" 
6 5 
6 9 

8 1 

9 6 



r 






4' 10" 

6 2 

7 6 

8 10 
10 8 



Abba 
Bbtwbbn 

TUBBB 



mSOQ 



I 



16.6 

21.86 

27. 

32.25 

89.25 



5 Q 



28.86 

29.1 

84.26 

89.5 

U.75 



^•8 



I 



,g| 



81.1 

86.85 

41.5 

46.75 

51.5 



1.984 
2.976 
5.952 
9.920 
14.880 



QQ 



884 

576 

1.152 

1.920 

2,880 



li 

r 



48 

72 

96 

120 

144 



In practice, economizers are frequently made 60 or more rows deep to obtain the 
heating surface. The over-all length of the economizer may be obtained for any even lumber 
of rows deep by allowing 7^ in. per row. 

Bzample. Suppose it be desired to install an economiser for each battery of 300 hp. boilers in 
the rear of the boilers, the width of the battery setting being 12^ 8". 

Based on 4 sq. ft. of economiser tube surface per b.hp.. there will be required 4 X 2 X 300 — 
2400 sq. ft. of tube surface per battery. If the economiser is made 10 tubes wide the heating surface 
per row b 120. the number of rows required is therefore 2400 / 120 a- 20; the lengtii of the economiser 
being 20 X 7^ - 146 in. or 12' 1". 



CHAPTER VII 

CHIMlfETS FOR POWER BOILERS 

THEORY, DESIGN AND CONSTRUCTION 

Draft Produced by a Stack or Chimney. The "head'' available, for oyerooming all frictional 
resistanoe to the gas flow and creating the final velocity of discharge when produced by a chimney, 
18 calculated by taking the diflference between the weight of the column of hot gases in the chimney 
and a column of the outside air of equivalent height. 

Let L » the height of top of chunney above the grate measured in ft. 
D » the diameter of chimney in ft. 
d IB density of the outside air (wt. per cu. ft.). 
dc » density of the chimney gases (wt. per cu. ft.). 
H a the total head produced by the chimn^ measured in inches of water. 
K » density of the water in the manometer. 
B 62.4 for a temperature of 70^ F. 
U 



^^-g(d-d,). 



12 



12 



Then H = LX — X(d-dJ 



5.2 



id - dj. 



From the law of perfect gases, P V ^ M RT, in which P is the absolute pressure in lb. per 
sq. ft. y s volume in cu. ft. M » weight in lb. iS » a constant «■ 53.35 for air. T >■ the 
abec^ute temperature degs. F. If F = 1, then M « the density or weight per cu. ft. P » 14.7 
X 144 - 2116.8 lb. per sq. ft. at sea level. 

Then If - d =-zr-- 

T 

If T » the temperature of the outside air and Tg « the temperature of the chimney gases, 

!_ ^ 39.7 
then dc * 



T,- 



.'. H - 7.64 L 



\T tJ 



(1) 



It is usual to take approximately 0.80, of the head H when the calculations have been based 
on the temperature of gases leaving the boiler to allow for the cooling effect of the flue and stack. 
The temperature of the flue gas may be approximated from the following table taken from a 
publication by the Qreen Fud Economizer Co, 

TABLE 1 



FfMnds wmtar evaporated from and at 212^ per 

■q. ft. heatinf surfaee per hour 

Square feet heating aiirface reqidred per hcnve- 

Batio of beating tognite' matam If Ji'aqHi. of 
G. S. la required per hp 

Probable tempcratore of ehimnoy gaaaa, degrees 
F 



2 


2.6 


8 


8.6 


4 


6 


6 


7 


8 


9 


17.8 


18.8 


11.6 


9.8 


8.6 


6.8 


6.8 


4.9 


4.8 


8.8 


62 


41.4 


84.6 


29.4 


26.8 


20.4 


17.4 


18.7 


12.9 


11.4 


460 


476 


600 


626 


660 


680 


660 


710 


770 


860 



10 

8.6 

10.6 

980 



159 



160 



POWER PLANTS AND REFRIGERATION 



The theoretical head produced for various heights of chimney and temperature of flue 
may be read direct from the accompanying diagram Fig. 1. 

Experiments conducted by Prcf, E. F, Miller at the Massachusetts Institute of Technology 
on the cooling of chimney gases have shown the f blowing results: 

TABLE 2 



SiM and Type of Stack 


Height 


TmpuuTUBaB, Dbgs. P. 


Initial 


Final 


Lon 


8' z 8' aquaTO bctok 


102 
100 
260 
260 


440 
406 
478 
860 


862 
829 
876 
816 


78 


V «)fftim4fr unllTMd atml 


76 


w <iVTT*«t^ ndiat hf^^ 


108 


W diaimrtm* ndiat brick 


46 







Let hg B loss of draft through grate, inches of watar. 
h^ « loss of draft throuc^ boiler, inches of water. 
hf » loss of draft throuc^ flue, inches of water. 
he a- loss of draft through chimney, inches of water. 

12 d y* 

— — — - tm the final velocity head inches of water. 
K 2g 

12 d y* 

Then 0.8 H - A. + A» + ^ + ^« + "^ TT* 

K, 2g 

Loss of Draft in Flues and Chhnneys or Stacks. The general form of the e]q>resBion giviiig 
the loss of head measured by the height of a column (in ft.) of the medium flowing in a pipe or 
duct is: 

LR y* LR 

*, -/ X -"T X — or/ X — r X velocity head, 
A Z g A 

in which h^ -> head lost measured in feet of air or gas column. 

L « length of pipe, duct or stack in ft. 
R ■» perimeter of pipe or duct in ft. 
LR ^ area of the rubbing surface in sq. ft. 
A » area of pipe or duct in sq. ft. 

V « velocity of flow (average over the cross-section) in ft. per sec 
/ B friction coefficient. 
g » 32.16 (acceleration due to gravity) ft. per sec* 

— — — the velocity head, in ft. 
2g 

If the head lost is stated in inches of water, then 

, h,K 



in which 



Then 



12 d' 

h « the head lost measured in inches of water column. 
K « density of the water corresponding to the temperature. 
d s density of the medium flowing corresponding to its temperature. 



h^f 



LR 



12 d 
A \2g K 

For a round section A » x D and A '^ Ht D^, 



f y« 12d\ 
\2g K ) 



(2) 



^UffUff 




Pio. 1. 



162 POWER PLANTS AND REFRIGERATION 

Then 

, ^4L / F« 12d\ ^ 

If it is desired to use the weight of the medium flowing rather than the velocity, the following 
substitutions may be made in (3) for round ducts: 

Q a Volume of flue gases, cu. ft. per sec. 
W » Weight of flue gases, lb. per sec. 

The weight of flue gas assumed in calculations is usually based on a fuel consumption of 
approximately 4 pounds per rated boiler horsepower at ncMrmal load and 20 pounds of air supplied 
per pound of coal.* The boilers are assumed to be operated at 50 per cent overload, in which event 
approximately 6 lb. of coal will be consumed for each rated boiler horsepower. The weight of 
flue gas per rated horsepower (not actual powor developed) will then be 6 X 20 or 120 lb. per hour. 

The theoretical and actual amount of air required for combustion is quite fully discniwad 
in the Chapter on " Fuels and Combustion." 



= — andF-- VV* '^ [tz) " 

d A Xd \Ad/ 



m 



The velocity head in inches of water 



(0.7854 D^Xd)* 
12XW* Xd W* 



2XgX 0.7854* xD*Xd»XK 204xX>*Xd 



Then A » / — - X ^, ,, ^ — 7 = 0.01967- r. 

D 204tXD*Xd D^Xd 

If the temperature of the water is 70^ in the manometer tube, then K «■ 62.4. 
d » 0.0749 corresponding to 70** F. for air. 
« .049 corresponding to 350^ F. for air. 
« .0414 corresponding to 600® F. for air. 
- .0393 corresponding to 550** F. for air. 
B .0375 corresponding to 600'' F. for air. 
Experiments by various authorities on the flow of air at practically atmospheric pressure in 
smooth sheet-steel ducts for velocities ranging from 15 to 30 ft. per sec. corresponding to the 
range of velocities as found in chimney and fan practice gave the following average friction co- 
efficient for use in the general formula given above: 

/ = 0.0035 for fi = 8 to 16 ft. corresponding to D » 2.5 to 5. 

As this value corresponds to smooth steel ducts it is advisable to increase this coefficient 
by 25 per cent for sheet-steel ducts and unlined steel stacks in practice to allow for rough joints 
and surfaces and to double the coefficient for brick or concrete surfaces. 
Then the practical values for use in the formula given are: 

/ = 0.0035 X 1.25 « 0.0044 for unlined steel chimneys or flues. ** 
= 0.0035 X 2. ■= 0.007 for brick chinmeys or flues. 

The head lost by friction, measured in inches of water, for the following stack temperatures 
using the above coefficients is: 
For Steel Stacks (unlined). 

L 

he = 0.00176 7- X TF» for 350* F., round sections. 



*Thm fud oonsumptkm per b.hp. b baaed on using a fuel having a ealorifie value of 18,600 Rt.u. per lb., aad aa 
fmt-tSi boiler, giate and fumaoa effldeney of 62 per cent. 

^ The coefficient reoonunended by aome authorities is approximately double the yalues gtvan bare. 



CHIMNEYS FOR POWER BOILERS 



163 



• W0I21 -r^ X TP for 500*" F., round sectioDs. 



« 0.0023 X — X TF* for 600** F., round sections. 



For Brick, Brick-Lined or Concrete Stacks. Multiply by 



0.007 
0.0044 



or 1.704. 



TABLE 3 

HEAD LOST IN UNUNED SHEET-STEEL STACKS 
Temp. 600<* 100 Ft. High 



Diameter 
Pt. 


he - 0.21 ^(inches water) 


Velocity Head Inches of Water 
TP 


1 


0.21 W* 
.0276 iy« 
.00066 W* 
.00213 TF» 
.000864 TP 
.0004 TT* 
.000205 TP 
.000114 W* 
.000067 W* 
.000042 TP 
.000027 TP 
.0000124 TP 
.0000064 TP 
.0000086 TP 
.0000021 TP 
.00000086 TP 


0.12 TP 


m 


.0287 TP 


2^.:.:::::.::::;:::.::::::::::.::::::::::: 


.0076 TP 


2H 


.00807 TP 


8 


.00148 TP 


SH 


.0008 TP 


4 ^ 


.00047 TP 


4H 


.000292 TP 


6^.:.:.:::;:::::::.:.::::::::::::::::::::: 


.00019 TP 


5H 


.00018 TP 


6 




.000098 TP 


7 


.00006 TP 


8 


.000029 TP 


9 


.000018 TP 


10 


.000012 TP 


12 


.0000068 TP 







The pressure loss and velocity pressure for various diameters of stacks may be read direct 
from the diagram, Fig. 2. 

For brick, brick-lined or concrete stacks, multiply the above values of h^ by 1.704; for any 
other heii^t multiply, by the height divided by 100. 

The draft losses through the grate and boiler may be approximated from the data given 
by the accompanying table. 

The following table is based on the values as read from chart, Fig. 3, in the Chapter on 
"Boikis/' etc.: 



TABLE 4 

INTENSITY OF DRAFT BETWEEN FURNACE AND ASH PIT TO BURN COAL 



Klndof Co«l 

« 



Combustion Rate "R" Lb. Dry 
Coal per Sq. Ft. Grate per Hour 



16 



20 



26 



80 



86 



40 



Force of Draft InebM Water 



m., Ind.. Kan. Bitaminoaa lO. 14 



Ala., Ky., Fa-« Tflnn. Bitominoua . 
M<L, Fa- Va- W. Va. Semi-Bitaminoua. 



Anthraoite Baekwbeat No. 1 



0.14 


0.20 


0.26 


0.88 


0.40 


0.48 


.16 


.28 


.81 


.40 


.49 


.60 


.18 


.26 


.86 


.46 


.67 


.71 


.80 


.46 


.64 


.88 


1.28 




.43 


.68 


1.00 


1.60 


■ • • • 


• • • 



46 



0.67 
.72 
.87 



164 POWER PLANTS AND REFRIGERATION 

He low of draft between the grale or rumoce and a point just beyond the damper boi of 
a boiler is about as follows when the boilers are operated at normal rati^. BitumiDOUt coal 
burned at the rate of 25 to 30 lb. per square foot of grate surface per hour. 

TABLE 5 

LOSS OF DRAFT IN BOILEBS 



BabMckftmiotB.. 




CHIMNEYS FOR POWER BOILERS 165 

The ]om of draft through the boiler will depend largely upon the method of baffling employed 
and inoeaaeB with the per cent rating at which the boiler is operated. The above figures should 
be inoreaaed by i^proximately 55 per cent when the boiler is operated at 150 per cent of its rated 
capacity, and by 75 per cent where it is run at 200 per cent rating. 

Velocity <d Gases thiough Flue and Chimney. The customary allowable velocities of gases 
in chimnejTB wbeai the design is based on 120 lb. of flue gas per hour per rated boiler horsepower 
varies from 17 ft. per sec. for a diameter of stack equal to 24" to 31 ft. per sec. for a 72" diameter 
and above. These figures correspond to a weight of 0.68 and 1.10 lb. per sq. ft. of area. The 
diagram (Fig. 1) gives the diameters ordinarily used for various amounts of coal burned per 
hour and oonesponding rated horsepower. The formula that is supposed to give the most eco- 
nomical diameter for an unlined steel chinmey or stack which is used by many engineers in 

this oountiy is 

d » 4.68 ^/(Hp,)^ 

in which d 'm the diameter in inches and Hp. is the rated capacity of the boilers served. The 
relation between the diameter and horsepower is given by the curve Fig. 1. 

The velocities in the flue breeching should ordinarily not exceed 90 per cent of the above 
vahies. 

The loss of draft in the flue may be approximated by means of the general formula (2) pre- 
viously given, using a coefficient of friction / ^ 0.0044 for imlined sheet-steel flues and / a> 0.007 
for brick-lined flues. The loss occasioned by right-angle turns and bends may be approximated 
from the data given in the Chapter on " Hot Blast Heating/' Volume I. 

Tlie following figures are frequently used by engineers for approximating the loss of draft in 
flues or breechings: 

Horiaontal flues, square or rectangular, 0.13 to 0.15 inches water per 100 feet. Increase these 
vahies by 50 per cent for brick-lined flues. Loss of draft, easy rig^t angle-bends, 0.05'' water. 

When eoonomiaers are to be installed, the temperature of the flue gas will be reduced to 
250* to 325^ and the total head ( H) should be calculated on a basis of these temperatures. 

The loos of draft through the economiaers should not be figured less than 0.3" water. 

When a superheater is used, allow approximatdy 0.15 inches additional draft loss. 

Tlie turns which the flue makes in leaving the damper box of the boiler, where it enters 
the main flue, and at the stack should be considered and allowed for. 

It is customary to make the flue or breeching approximately 10 to 15 per cent greater in area 
than the area at the top of the stack to which it connects, the cross-section being reduced in 
proportion to the volume of gas to be handled as the flue passes the boilers in succession. One 
prominent boiler numufacturer recommends a flue area of 35 sq. ft. per 1000 b Jip. 

The following chimney formula by WiUiam Kent is largely used by engineers in this countiy . 

The formula is based oA the following assumption : 

The friction head in the chimney is considered as equivalent to a diminution of the area by 
an amount equal to lining of inert gas 2" in thickness. 

If A ^ actual area, sq. ft. 

E « effective area, sq. ft. 
D a diameter, ft. 
Then ^ - A - 0.60 VI7 

The draft power of a chimney varies directly as the effective area E and as the square root 
of the hei^t B. 

The formula tor horsepower of a chimney will take the form Hp. « C J? V B, in which C 
is a constant. The vahie of C. as obtained by WiUiam Kent from an examination of a large niunber 
of chimneys is 3.33 when 5 lb. of coal is burned per boiler horsepower per hour. 



166 POWER PLANTS AND REFRIGERATION 

The formula for the horsepower rating of a chimney is then 

Hp. = 3.33 E VT = 3.33 (A - 0.6 V^) V"b 

0.3 Hp, 
or ^ « — , — . 

TABLE 6 

SIZE OF CHIMNEYS FOR STEAM BOILERS 

Kent's Formula 





Area 

(A) 

Sq.Ft. 




Height of Chimney 


Equivsleot 


Diam. 
Inches 


60 
Ft. 


60 
Ft. 


70 
Ft. 


80 
Ft. 


90 
Ft. 


100 

Ft. 


110 
FL 


126 
Ft. 


150 
Ft. 


175 
FL 


200 
Ft. 


226 
Ft. 


250 

Ft. 


800 
Ft. 


Sq. Chim- 
ney Side of 
SqVE+4 
Inches 




Commercial Horsepower of Boiler* 


18 


1.77 
2.41 
3.14 
8.98 

4.91 
6.94 
7.07 
8.30 

9.62 
12.67 
16.90 
19.64 

23.76 
28.27 
83.18 
88.48 

44.18 
60.27 
66.75 
63.62 

70.88 

78.64 

96.08 

113.10 


0.97 
1.47 
2.08 
2.78 

8.68 

4.48 
6.47 
6.67 

7.76 
10.44 
18.61 
16.98 

20.88 
26.08 
29.78 
84.76 

40.19 
46.01 
62.23 
58.88 

66.88 

73.22 

89.18 

106.72 


23 
86 
49 
66 

84 


26 
38 
54 

72 

92 
116 
141 


27 
41 
68 
78 

100 
126 
152 
183 

216 


29 
44 

62 
88 

107 
183 
163 
196 

231 
811 




















16 


21 


















t 


19 


24 


66 
88 

113 
141 
173 
208 

245 
380 
427 
636 






1 




!....i...: ■ 1 


22 


27 








......... ....|. ........ 




24 


80 
88 


119 
149 
182 
219 

258 
848 
449 
566 

694 
886 


• •• • • 

166 
191 
229 

271 
365 
472 
693 

728 

876 

1,088 

1,214 


..... 






■ • * • • 


. . 




• • • • - 


27 
80 


86 


204 
245 

289 
889 
508 
632 

776 
934 

1.107 
14i94 

1,496 
1,712 
1,944 
2,090 














82 


89 














35 


42 






316 
426 
551 
692 

849 
1.023 
1,212 
1,418 

1,639 
1,876 
2,180 
2,899 

2,685 
2,986 
8,637 
4,362 












38 


48 
















43 


64 








596 

748 

918 
1.105 
1,810 
1,631 

1,770 
2,027 
2300 
2,692 

2,900 
8.226 
8,929 
4,701 










48 


60 


















64 


66 










981 
1,181 
1,400 
1,637 

1,898 
2,167 
2,469 

2,771 

3.100 
3,448 
4,200 
5,026 








59 


72 












1,253 
1,485 
1,736 

2.008 
2,298 
2,609 
2,989 

3,288 
8,657 
4,455 
5331 


i.666 
1380 

2,116 
2.423 
2,760 
3,098 

3.466 
3,855 
4.696 
5,618 


• • • • • 

2,065 

2318 
2,654 
3,012 
3393 

8,797 
4323 
5,144 
6,155 


64 


78 








1 . 


70 


84 






::;;; .:..!:.:: 




75 


90 














80 


96 








. . |- -^ • -^ 




86 


102 




[ 


1 






91 


108 








1 






96 


114 








1 






101 


120 




t 


1 






107 


182 






i 






117 


144 








( 






128 










, 















* Based on a consumption of 5 lb. of fuel per boiler horsepower. For any qther rate multii^y the tabular 
figure by the ratio of 5 to the mayimum expected coal consumption per horsepower per hour. 



Table 6 is based on the above formula. The B, <Sk W, Co. recommend that when the fuel 
used is low-grade bituminous of the middle or western states that the sizes given be increased 
from 25 to 60 per cent, depending upon the nature of the coal and capacity desired. If the gas 
makes more than two turns it is advisable to increase the diameter as given by the table by one 
size. The height must be increased at least 30 per cent if economizers are to be used. 

The above table may be applied to heating boilers, the equivalent rating in square feet of 
direct radiation being approximately equal to hp. rating X 100. 

Example. The method of procedure in determining the dimensions of a flue or breeching and a 
brick chimney will be explained by the following example. The layout of the plant is shown by Fig. 4. 
There are three 150 hp. return tubular boilers to be served; the grate surface of each boiler is 30 sq. ft. 
as given by the table of dimensions for return tubular boilers. The maximum weight of flue gas per 



CHIMNEYS FOR POWER BOILERS 167 

hour per bmler horaepowcr will be asnimod sa 120 lb. Tho fluo has two riKhtanele turna upon enter- 
ing the Sue and stack. The mesaured length of the flue is approximately 40 ft. The fuel awnuned it 
Pemuylvuiim bituminouB and the total grate area is 90 aq. ft. If 6 lb. of coal per boiler horsepower is 
anuned for the fuel eonoumption, then 3 X 150 X 6 / 90 — 25 lb. per sq. ft. per hour is the rate of 
Oombuati»n R. 

The looBof draft through the grate from the Hiagrain Fig, 1, in the Chapter on "Power Boilers," 
for this late of combustion ia Aj •■ 0.31 inches water. The loss of diaft tbrou^ the boiler ht ~ 0.30 



Fia. 4, Latoct or Phixr. 



hf - 0.15 X TZ; + 2 X 0.05 - 0.16 inohe*. 



168 



POWER PLANTS AND REFRIGERATION 



■ height ot 110 tt, the looa from the diacram Fie- 2 per 100 ft. la 0.0256" and the veloeitjr head ii 
0.067". 

The Ion through a bnck stack £■ 1.10 X 0.0256 X 1.7 - 0.048". 

Then 0.8 H - 0.31 + 0.16 + 0,048 + 0.067 - 0.685" 
or H -OIZ". - '' '/.'^ 

The total theoretical head produced by a stack having the above duDenaioiu for a temperature of 
600" F. ia given by the diagram, Fig. 1, aa H - 0.75". This is seen lo be auffideDl Hdsht ot chirooey 
above foundation will be made 115 ft. 

The standard dimensjona for a radial brick stack of thia siie will be found in Table 14. The aira 
of thia stack is 16.S sq. ft. The area of the flue ftDm the last boiler in the line if made 10 per cent greater 



TABLE 7 
DIMENSIONS OF BREECHING CFIG. V) 





BO 


M 


TO 


76 


10 


» 


OS 


1.0 


1(5 


... 


155 


170 


19. 


BoOn- IMameter. Inches 


4S 


M 


60 


60 


80 


60 


66 


66 


72 


71 


72 


78 


78 


\ 


"W^.t-'K*." 


lit 


18 
72 


IS 
S6 
U 


20 

to 

60 


EO 
40 

: 


EO 
60 


a) 

40 
60 
80 


24 
48 

7S 


24 
72 












i: 
a 


Hdfht of biwebinc at 
Hri(bt of bnechinc at 


M 
4th 










It 

f 


Heiitat of braecfalnt Bt 

H^r'o™b^tai' U 

boilv, am. line, in 

Hrijht of b«cUnE U 


Ai 

'3d 

ith 








.. 






19 H 
S9 

58 W 
78 


19 M 
89 

MM 
78 


SS 

79M 


26 N 
51 

79 M 
10« 


MM 
53 

loe 


80 
60 
90 
120 


80 

90 
120 



in area will be 17.5 sq. ft. Tho width of opening at the base of stack should not exceed 33 per cent of 
the outside diameter of chimney. 

The outwde diameter at bottom for this siie chimney from Table 14 is 10.4 ft. The width of 
flue will b* made 3 ft.; the height of Sue at chimney will be 17.5/3 = 5.S3 ft or Q'-IO", 



CHIMNEYS FOR POWER BOILERS 



16d 



C hium c y for Tall Office and Loft Buildings. The chimney car stack f (ht a tall building is 
a Bpecwl case in which the height is frequently fixed by the hei^t of the structure or adjoining 
btiikiingH. In this case a diameter is assumed and the preceding method applied. 

Smoke Breechings. Smoke breechings are ordinarily constructed of steel plate and are 
made 10 to 15 per cent greater in area than the area of Uie chimney with which the breeching 




FlO. 6. BOUMD Stbkl BBUCmMO. 

oonneets. The smoke connection fear each boiler should be provided with a damper that may 
be operated from the floor either by a damper regulator car adjusted by hand. 

Breediing for 50 horsepower bcHlers and smaUer are ordinarily made of No. 14 steel; No. 12 
■teel for boilerB 60 to 115 hp., and No. 10 steel fcHr 125 hp. and larger. 



TABLES 

WEIGHTS OF ROUND BREECHINGS (FIG. 6) 



1 


^1 








WBOBTB 




• 


4 


Dumm 
B»— mmo. 












"Sj 
















«1 


iNCBBi 




Two BoOan 






Threa BoOan 




l« 


Two 


ThrM 




















s 


BoOan 


BoOan 


No. 16 


No. 14 


No. 12 


No. 10 


No. 8 


No. 14 


No. 12 


No. 10 


No. 8 


S6 


68 


88- 


84 


800 


860 


460 


600 


760 


600 


760 


1000 


1260 


4S 


•8 


88 


40 


400 


460 


600 


760 


900 


760 


1000 


1260 


1660 


48 


78 


86 


42 


460 


660 


700 


860 


1100 


960 


1200 


1460 


1860 


64 


84 


40 


48 




660 


900 


1160 


1400 


1100 


1600 


1960 


2400 


60 


90 


44 


68 




800 


1100 


1400 


1700 


1860 


1860 


2400 


2900 


66 


96 


60 


60 




900 


1260 


1600 


1900 


1600 


2100 


2700 


8800 


72 


106 


68 


64 




1860 


1900 


2400 


2900 


2800 


8800 


4000 


4900 


78 


lis 


66 


68 




• • • • 


2400 


8000 


8700 


. • • . 


4000 


6000 


6200 


84 


UB 


60 


72 




• • • • 


2700 


8400 


4100 


« • • • 


4600 


6700 


6900 



STEEL CHIMNEYS 

Oiqred Sted Stacks or Chimneys. Steel stacks that do not exceed 75 feet in height (ht mofe 
than 4 feet in diameter are <»dinarily guyed by 4 steel cables to resist the wind pressure. The 
gojTB are attached at approximately H oi the height. As they do not depend upon any foun- 
diUion fcHT stability, they are frequently supported directly on the boiler breeching. 

Sdf-Siqiporting Steel-Plate Chimneys. This type of chimney dq)ends upon the weight of 
the foundation, ^diidi is almost invariably constructed of concrete, for its stability. Steel ohim- 
ne3rs are ordinarily lined with A^i" fire-brick for a height of 25 to 50 ft. above the breeching 
connection, the remaining pcnrtion of the lining being constructed of common brick. 

A common type of self-suppcnrting chimney and details of c(mstruction are shown by Fig. 7. 
Hie letters refer to the data given by Tables 10 and 11. 

Thickness cf Plates. The thickness of steel plate required (or any section is determined by 
treating the shaft as a uniformly loaded cantilever beam, the total uniformly distributed load 
being equal to the product of the unit wind pressure per sq. ft. and the projected area of the 
shaft for the porticm above the secticHi or point being considered. 



POWER PLANTS AND REFRIGERATION 



CHIMNEYS FOR POWER BOILERS 171 

-- 1 /2d4-di\ 

The moment ann y of the horisontal wind force R is equal to — - « ( --; -;- J in which z is 

3 \ a -f- oi / 

the distance down from the top of chimney to the section considered, d = outside diameter at 
the top and di = outside diameter at the section considered. 

The formula for determining the stress in the plates at any section is the same as given by 
Fig. 12 for radial-brick chimneys in which di » the outside diameter of the chimney at the section 
considered in inches, dt = the inside diameter in inches for the same section. The maximum 

W M 

stress occurs on the leeward side, /i = "7 + "t:* The allowable stress, lb. per sq. in., should not 

A S 

ordinarily exceed 6000 lb. for single-riveted joints and 8000 lb. for double-riveted joints. The 
least thickness that should be used is ^/le in., and where stacks are over 7 ft. in diameter, or 150 
ft. high, the thickness should not be less than K in* 

The following simplified fcnmula, for all practical purposes of design, may be used in place 
of the more complicated formula for cylindrical chimneys: 

t » thickness of shell, inches. 

h » distance from top to the section under consideration, feet. 

r = radius of section under consideration, inches. 

t = ■ for single-riveted girth joints, umt stress 6000 lb. per sq. in. 

' " ^^^^^ for double-riveted girth joints, umt stress 8000 lb. per sq. in. 
1000 r 

The above formula is based on a wind pressure of 25 lb. per sq. ft. of projected area. 

Example. Determine the stress in the plates for the section / at top of beU-shaped section \b'-(y' 
above top of foundation. Fig. 7 for a steel chimney B -■ ISC high and A -> 66" diam. (Reference 
No. 24, Table 10.) The thickness of plate for the first section K is 3/8", the inside diam. of chimney 
/ « 8^-4", or 100", the diameter at the top / -■ 6'-4". Assume unit wind pressure p -> 25 lb. per 

(6.33 + 8.33\ 
9 / 

1 /2 X 6.33+833\ 

-990 sq.ft. i2-990X25 - 24.760 lb. horisontal wind force, y ---(150 -16) I ^ ^^ . ^^ — I 

3 \ D.oo + 833 / 

- 64 ft. Wind moment M ^Ry » 24,760 X 64 X 12 »- 19,008,000 in.-lb. The section modulus 
of the chinmey at section / is: 

„ 0.098 (l00.76 - 100 ) 

iS -— -^ 2960. 

100.75 

135 
The weight of the sted shaft above / is approximately — X 30 - 27 tons or 54,000 lb. 

XOv 

The area of steel at section / is: 

A - 0.7854 (100.75 -Too ) - 118 sq. in.. 

W 54,000 M 19,008,000 

/i — — — — 460 lb. per sq. in. stress due to weight of shell, /i " "T ■■ — zrir — ■■ 6443 lb. 

JL 1 lo a ^vou 

per sq. in. stress due to wind moment. Total stress // « /i + /s "* 6903 lb. per sq. in. 

Seams are ordinarily single-riveted except for beU at base of chimney. 

The rivet spacing should not be more than 16 times the thickness of plate, or more than 
6 inches. 

Foundation Bolts for Self-Supporting Steel Chimneys. It is quite generally assumed, al- 
thoug}i oth^ analyses are sometimes made, that the chimney is fixed at the base^ the neutral 






FOUHDATION fUN 

Flo. 7. Sn^-SuTPoirnNo Stbei. Staok. 



CHIMNEYS FOR POWER BOILERS 



173 



S, 



1^ 









PV|»I»A 






io9<isy»ii 










Is 



Hi 




|0 4i|>Pii 






><-<«>Aei»-ri< 



IPlPllIlilllPIlIillll^^ 






^88! 

ft % * 

icocf( 



_.S8SgS4?JS* 



:SI 


















MMMMMMMMMMMMMMMMMMMMMMMMMMMMMMMMMMMMMMMMMMMMM 

ag8ag8ag8g8SgaSg8a^89^89899989SiSggiStg9g89ggggg 



P8 



sgs;sgssgS58 sagsags;gaagggs8S3ga38 3gss!g2iSggg8g8g 



7 

I** 
PS 



SgsssgSiSgssgSiggSigg&agaggagag^ggsagggiSggiSggggggg 






fe85sg&sgsasagsa8sagag8 a2as3ga3agggisggig8ggagag 






OOMOOCOOOCOMOOflO 






^ JilMUIIQ 















M. 



^•a*S5^;:S*SJS«S3S*SS2«SS«2gSg$S$gg538giSS§g?iS8g8g| 



""^'•s^'i 






^ggg$gg$gS!S^g$ggg^!$gg!9gg8gS8gS^gSgggggggSgSgg 

V^ ^^^ ^^^ rw ^^ ^*^ ^^^ ^^^ r^ 1^^ 



»!:« 



•^^^^^ 



XrS! 



:5!:j::J!:f:«:« 



xj::« 















22222222222222^2^2^2g22s32gagS^S3SSg»^SSS^ggS^aS^ag 



gSSg99;(S3?ga|22§gggS||g3|2§|||2|||||g^||P|| 






•JBOH Jw (loo 













y 

ran 
JH5L 



^g2^g2^g2S2gS2ggSg8^SS^S§S^Sg^gg§$g|gg§g|g§i§ 



ksS8SSSSS99939S223SSSSSS^^^SSS333SSSS33SSSSSS 



^ v4 iH ^ vH ^ ^ ^ ^ M M « 09 « ^ 9 ■« »o 2 « K- r* « A A q^iHi iH e|e^^ia b;a2^i-4 ^^^ 



XlMlUipiQ 



-*«*»''*»^**23aassafcsa8aaasiS8saa$saasissss$sssa55!5 



"•M 



POWER PLANTS AND REFRIGERATION 



CHIMNEYS FOR POWER BOILERS 



176 



axis being taken on the center line x — x, Fig. 8. In this case any bolt will be stressed in propor- 
tioii to ito disianoe frc»n the neutral axis, the bolts on the windward side of the neutral axis 
being in tension. 



I«t dgf d^ di, 

di^+d^+d^ + 



(2« = distance of bolt center from neutral axis, feet. 
J.« - 2d«. 
r » radius of bdt circle, feet. 
N = number of bolts. 

Zd* 
Q « — section modulus of anchor bolt group. 

S i- stress in extreme bolt, lb. 
M » wind moment, ft.-Ib. 




etrelM 



m in which 26 is the assumed wind pressure 

in lb. per sq. ft. of projected area. 
D "■ diameter of cylindrical stack (ft.) and H » height (ft.). 
R ■■ righting moment. 

— weight of chimney X r. 
P B overturning moment, ft.-]b. 

« M - «. 
A = area of bolt at root of thread, sq. in. 
/ « allowable fiber stress, lb. per sq. in. Ordinarily limited 
to approximately 12,000 lb. per sq. in. 

il--eq.in. 

Hie foundatioQ is designed in the same manner as for a brick chimney, the soil pressurs 
being ordinarily fimited to 2 to 3 tons per sq. ft. See Fig. 12 for formula. 
*A formula fraqiuBtly Hied by anginem la. 



8 - 



2Af W 

tN N* 



176 



POWER PLANTS AND REFRIGERATION 



Example. Determine tlie total stress S and unit stress / for the anchor bolts for the chimney 
given in the preceding example. H -> 15(K; D -> 7.5' (approximate average) ; r -> 7.1'; N » 10. Diam. 

25 X 7 5 X 150* 
bolts 2yi'\ A -= 3.7 sq. in. if - — ' - 2.109.375 ft.4b. R " 70.000 X 7.1 - 497.000 



242 
ft-lb. P - If -* iJ - 1.612,375 ft.-lb. 2d«-242. Q - — - 34. 

f . X 

47 425 
/ - ' - 12.820 lb. per sq. in. 
3.7 



1,612,375 
34 



- 47,425 lb. 



TABLE 12 
BASIS OF SELF-SUPPORTING STEEL CHIMNEY ESTIMATES (For Tablfls 10 or 11) 

RiDBISTANCB TO WIND PRBBSURB, 26 Lb. PBR SQ. Ft. FACTOR SaFBTY, 9. COAL CONSUIfPTIOM, 6 LB. BlTUMIMOOa 

CkiAL PBR Hp. Hr. Draft Pressure Barombter, 14.7 Lb. OuTsmB Am at 60** F. and Gasbs in CHnoonr 
AT 60(P F. Gratb Surpacb Having 40% Efpbctivb Arba op Opening. Connecting Flues Dirbct wns 
20% Greater Arba Than Chimnby Area. Top op Chdinby Extended Above Surroundinqb Which 
Would Divert Dirbct Outsidb Air Currbntb. Proportionate Loss in Draft Power bt Hobibontai. 
Connecting Flues. 



Longth of flue in ft. . . 
Per oent Iom In drmft. 



60 
1.0 



100 
4.8 



200 
18.6 



400 
26.2 



600 
36.6 



1000 1600 
43.9 49.6 



2000 
58.7 



3000 
67.8 



Wmter-heating eeonomiaav placed in fhies 
drmft 20 to 60% proportionate to tampermtara 
and area reducuon. 



BRICK CHIMNEYS 

Ordinary Brick Chimneys. Chimneys built of ordinary brick have a batter of Vis" to yi" 
per foot on each side; the diameter at the base is ordinarily made about VlO of the height. 
The top is protected by a cast-iron cap or the ornamental part at the top laid in Portland 
cement lime mortar (1-1-4 mix) and capped off with the same material 

The following tables may be used in the preliminary determinations of the thickness of the 
walls required for brick chimneys, but should always be checked for stability by the method as 
outlined later. 

The brick for the external walls should be selected hard burned and laid in cement lime mortar, 
1-2-6 mix for the upper part and 1-2}^S mix for the lower part. 

The core may be second class for brick laid in lime mortar, no cement being used. 

(Extract from " The Locomotive.") 

The core is by many engineers extended up from the base of the chimney 25 to 50 feet, 
but better practice is to run it up the whole height, stopping it off 8 or 12 inches from the top 
and not contract the outer shell. Under no circumstances should the core at the upper end be 
built into or connected to the outer stack. This has been done in several instances and the 
result has been the expansion of the inner core, which lifted the top of the outer stack squarely 
up and cracked the brickwork. 

Radial-Brick Chimneys. This type of chimney is built of moulded radial perforated brick 
which conforms to the curvature of both the inside and outside of the chinmey. The standard 
Cmtodis brick are made 4J^" thick and 6J^" wide, the radial lengths being 4", 6H", 7H", ^H" 
and 10^". The perforations are 1" square; there being 6 in the smallest and 15 in the largest 
size brick. The dead air spaces provide an exceUent insulation, tending to keep the gases hot 
and thereby promoting good draft. The perforations aid in securing a uniform quality during 
the burning period and also serve to increase the bonding action between the mortar and brick. 
The usual taper or batter of radial-brick chimneys is 1.8 to 2 ft. per 100 ft. 

These chimneys frequently set on an octagonal base of selected common red brick. The 
height of the base is made approximately 0.14 of the total height of chimney above the foundation. 
These chinmeys are usuaUy only lined to a height of approximately 35 to 50 ft. above the 
flue connection, thickness of lining 4)^", air space 2". 



CHniNEVB FOR POWER BOILERS 





Tmaoaai or Wallh — iNCaaa 










Om-ETtat 
tTop 


Thlcknwot 
Conor 

tiBlBflllclMi 




1 


» 


12 

I 


Irt 20 lot from top— 4' 






aisSil^ 






i<«>od»£S: 





















The Kdlogg Co.'s eanugfited perf(»«t«(l r&diftl brick is ahown hy Fig. 9. Hie differeDt wall 
thkJmesBM are obtained by a combination of theee bricks. They ore manufactured in maoy dif- 
forat radii eo that when laid in the wall they will form the normal ring with thin, even joiota. 

The method of bonding is shown by Fig. 10. It will be aoticed that there is a perfect inter- 
loddng of tbe bridu. The fact th«t the bricks are pofectly bonded, added to the fact that 
each brick is keyed to the other through tbe mortar in the perforations, makqa a remarkably 
straig wan to radst beat Btraine. Hg. 11 giTes the thickness of Kellogg radial-txick chinmeys. 



178 



POWER PLANTS AND BEFRIGERATION 



The dimeiuioD across the flats for an octagonal base may be determined as indicated b; the 
roUowing example. Fig. 14. Hie eiie across the flats for a ISO* x 7'-6" chimnej', the heigbt 
of the octagonal base being 30', first find the diameter at the bottom of column or 150' dawn 




no, 0. Pektokited 



71— 



28-* U- 



Fn. 10, MnsOD or BoHmxa Rauiai. Brick Flo. ] I . HmaET 

Waua of RadiUcBiuitk Okunktb. 

from top from Table 14, or 14'-5" in thia case. Continue the taper of the IfiO" chimney to ISC'. 
In 160' the outside diameter has increased from S.75' at the t(^ to 14.42' at the bottom at £.87' 

— ocntiDued to ISC this increase would be 5.67 X —wfi.Sl'.which.BddedtoS.TS'.glvM 16JS6'. 



CHIMNEYS FOR POWER BOILERS 



179 



Hie acro8B flats dimension of the base should be the next larger even inch or 15'-7''. The base 
18 made circular inside with an c^set of 1 inch from the column above. 

TABLE 15 

WIDTH OF FOUNDATIONS AT BASE FOR RADIAL-BRICK CHIMNEYS* 



Hdght Chimney 






INSIDB DlAMBTKR AT TOP — FBBT 






in Feet 


8 


4 


5 


6 


7 


8 


9 


10 


90 


iV-i 
12 -< 

14-4 
16-1 

• • • 


r i2'-o" 

\ 18-0 
( 14-41 
( 16-41 

; IBS 

17-9 
19-0 
20-« 
22-0 


18'-0" 

14-0 

16-0 

16-0 

17 -« 

18-6 

19-9 

21-0 

22 -« 


18'-9" 

14-8 

16-6 

16 -« 

17-8 

18-10 

20-0 

21 -« 

28-0 

• • • • 

• • • • 

• • • • 

• • • • 

• • • • 


• • • • 

16'-6^ 

16-6 

17-« 

18-8 

19-9 

21-0 

22 -« 

28-9 

25 -« 

27-0 

28-6 

• • • • 


• • • • 

16'-0'' 

17-0 

18-0 

19-8 

20-41 

21 -« 

28-0 

24-41 

26-0 

27 -« 

29-0 

• • • • 


• • • • 

• • • • 

• • • • 

• • • • 

20'-0" 

21-0 

22-8 

28 -« 

25-0 

26 -« 

28-8 

29-9 

81-6 

88-0 




too 




110 




120 




180 


21 '-41* 


140 


22-8 


150 


28-6 


160 


24-9 


170 


26-8 


180 


27-6 


190 


29-8 


200 


80-9 


210 


82-6 


220 


84-8 











* VelneB intefpoleted from cnrveB by M. IF. Kettogg Co, Maximum unit soil prenure at outer edge of foundatkn 
due to deed and wind loads does not ezoeed 2 tons per square foot. 

The width at top of foundation is made approximately I'-d" wider than the outside diameter 
at base of stack, as given by Table 14. 

The width of base required may be checked by means of the formula, Fig. 12. 



TABLE 16 

SAFE BEARING POWER OF SOILS 
Ira O. Bilker 



Kind of Material 



Rode, the hardeat* in thick layers, in native bed 

Rock, equal to the beet aahlar masonry 

Rode equal to the beet briek masonry 

Roek, equal to poor bride masonry 

Clay in thidc beds, always dry 

Clay in thidc beds, moderate^ dry 

Clay, soft 

Gravel and coarse sand, weU cemented 

Sand, dry, c o mpact and well cemented 

Sand, dean dry 

Qnidnmnd, aUuvial aofla, etc 



Safb Bearing Power 
Tons per Sq. Ft. 



Minimum 



200 
25 
16 

6 

6 

4 

1 

8 

4 

2 

0.6 



Maximum 



80 

20 

10 

8 

6 

2 

10 

6 

4 

1 



Average 



27.6 
17.6 

7.6 

7. 

6. 

1.6 

9. 

6. 

8. 

0.76 



In his book, ''Allowable Pressures on Deep Foundations/' Elmer C, Corthell gives the following 
summary: 

Tlie pressures of stable structures on fine sand range from 2.25 to 5.80 tons, average 4.5 tons. 

On coarse sand and gravel, 2.4 to 7.75 tons, average 5.1 tons. 

Sand and clay, 2.5 to 8.5, average 4.9 tons. 

Alluvium and silt, 1.5 to 6.2, average 2.9 tons. 

Hard clay, 2.0 to 8.0, average 5.08 tons. 

Hardpan, 3.0 to 12, average 8.7 tons. 

Clay, 4.5 to 5.6, average 5.2 tons. 



180 



POWER PLANTS AND REFRIGERATION 



TABLE 17 

CARRYING CAPACITY OF VARIOUS TYPES OF PILES FOR AVERAGE SOIL CONDITIONS 



• 

Siaaof Pile 


Surface 
Area, 


Frictional 

Carrying 

Capadty at 

800Lb. 
perSq. Ft. 


Bearing 

Area 

at Foot or 

Point, 
Square Feet 


Direct 

Bearing 

Capacity 

at 6 Tona 

perSq. Ft. 


Total 
Carrying 


Wooden Pile 80 ft. lone. 
Diameten 12" and r' 


74.6 
94.8 
110.0 
126.7 
188.6 
188.6 


Tona 
11.2 
14.2 
16.5 
18.8 
20.0 
20.0 


0.270 
.205 
1.07 
1.895 
1.58 
7.10 


Tona 
1.85 
1.08 
5.85 
6.96 
7.90 

85.5 


Tona 
12.6 


Conereta Plla 80 ft. lone. 
Dlametera 18" and 6" 


15.2 


Concrete Pila 80 ft. tons. 
Diameters 14" and 14'' 


21.9 


Conerete Pile 80 ft. lone. 
Diameteis 16" andl6" 


25.8 


Conerete Pile 80 ft. long. 
Diameten 17" and 17" 


27.9 


Conerete Pedestal Pile 80 ft. long. 
Diameten 17" and 8 ft 


55.5 







NoiB. — Ordinarfly it takes about twice the number of pflee for a diimney foundation that would be required 
for the dead load on^. Conerete piles coat about |1.50 per foot in place. 

TABLE 18 
NORMAL TOTAL DEPTH OF FOUNDATION FOR RADIAL-BRICK CHIMNEYS* 



Sixeof Chimney 



76' X 
100 z 
126 z 

126 X 
160 X 
160 X 
176 X 
176 X 10 
200 z 7 
200 X 10 

K 9 



8' op to and including 100' z 6'.. 

7 up to and indudfag 100 z 8 . . 

8 up to and including 125 z 8. . , 
8-6" up to and including 125 z 10. . , 
4 up to and induding 150 z 8. . , 
8-6 up to and induding 150 z 10. . , 
4 up to and induding 175 z 9. . , 



up to and induding 200 z 9. 
up to aiul induding 226 z 16, 



Total 


Depth of 


Foundation 


4'-6" 


5-0 


5-0 


6-0 


6-0 


6-6 


6-9 


7-0 


8-0 


9-0 


10-0 



* The valuee given by table are mimimum and must frequency be considerably increased to suit local conditioa% 
for example adjoining building excavations and foundations that are in ezeess of the depths stated. Foundations, 
reinforeed. should be stepped off as shown by Fi0kl8 and 14. 

TABLE 19 

DEAD LOAD OF RADIAL-BRICK CHIMNEYS— IN TONS OF 2,000 POUNDS* 









INSIDB DlAIOTBI AT TOP, FOBT 




^ 


Height 


















fSk 




















8 


4 


5 


6 


7 


8 


9 


10 


90 


9( 


\ 98 


110 


122 


• • • 


• V • 


• • • 


• • • 


100 1 


1( 


1 120 


181 


148 


161 


180 


• • • 


• • • 


110 ] 


3i 


1 148 


155 


167 


188 


206 


• • • 


• • • 


120 : 


m 


> 170 


185 


198 


218 


287 


• • • 


■ • • 


180 




202 


218 


231 


252 


278 


296 


818 


140 




287 


266 


270 


290 


810 


837 


867 


160 




277 


296 


811 


830 


858 


876 


400 


160 




817 


840 


867 


875 


402 


428 


447 


170 




862 


888 


410 


425 


454 


475 


500 


180 






• • • 




480 


510 


637 


566 


190 






• • • 




685 


570 


692 


617 


200 






• • • 




600 


632 


657 


686 


210 






• • • 




• • • 


• • • 


727 


760 


220 






* • • 


... 


• • • 


• • • 


804 


848 



* Values interpolated from curves by M. IT. Kettogg Co,, and are for round radial-hrick chinmeys excluaive of the 
wdi^t of founda t io n . 



R POWEIR BOILERB 



181 



Cliiinii«7 Deiixn Foismla: 

Let P — hoiiioDtiJ wind preesura, lb. per sq. 
ft., ordinarily assumod as 26 lb. 
p«r eq. ft. for tound chiouug^ 
« — z — any section distant > from top of 



( (^ j - iffojected 

£ vIioriKinta 



area above z — ; 
V IioriKintal wind load, lb. 



1/ — diitance from z — z la oeitter g t avily 
of portioa aboTB x — x. 
m wind moment, ft.-lb. 




- 0.7S64 (A' - <kO. 
8 — Section tnodutus. 
0.0082 <di* - A*> 
4 
W -• wtj^t of ohimney abore z — 
/. - 

/. - 



ward aide, tone per tq. ft due to IF. 
mpnMve itrev at edge on 1m- 
WBid aide, tone per eg. ft due to J/. 



Wtodwari tfdo/. - ^ - ^. 



+ i iiiiiiiiiiiiiiiiiiiiiiiiiiiiiiiiii^ 

- _ ^^fflmffllHHI Ilir *^">-S 

♦pj pipiiiii"''^! -f- f 



AtH/Uiri Sf'l Prmut^ 

i—iS " " " " " " t"^ 

FlO. U. 



Compreeeion. 
/s and /( abould not exceed tbe following valuaa, 
toDB per eq. ft for radial-brick ctunmeya: 

Jf ozimutn T«nnon Ifaxtmum Comprauion 

Below IGO'.... 2to2H 200' and bolow 10 

150 to 200*. . . . 4 to IH Above 200' 21 

Above 200'.... 

FmrndaHont. Calculate wind moment. Mi for 
ehimney above ground line. 



Af, -PA, 



■(^> 



A — ~ ~ BBCtion modulus bate. 

TT) V combined wel^t of chimney and 
foundstion. 
Keanttaot Soil Piaianra. Tons per aq. ft. 

Pi - — — coiiipt«e«ion per sq. ft due to W\. 

Pt — — — oomprenion per sq. ft due to tb* 

'" '^' ' Momeot, ttu 



182 POWER PLANTS AND REFRIGERATION 

The weight of 4^" brick lining may be approximated by the formula, weight in tons is equ^ 
to height of lining in feet wnes outside diameter in feet taken at the midpoint times 0.063. . 
Plain concrete is figured as weighing 1.9 tons per cu. yd. and if reinforced, 2 tons per cu. yd. 

Example. Determine maximum compression, tons per sq. ft. at the base of column, for the chimney 
shown by Fig. 13 and also the maximum soil pressure, tons per sq. ft. Assumed wind pressure 25 lb. 
per sq. ft. See formulae Fig. 12. 

Area of section at base A - 0.7854 (16« - 12.3«) « 80.9 sq. ft. 

, « 0.0982 (16< — 12.3*) ^^^ 

Section modulus at base 5 «= — ■* 267. 

16 

The total weight of brick column from Table 19 is TT » 495 tons (interpolated). 

8'_9'' + 16'-0" 
The projected area of column is, X 180 - 2228 sq. ft. 

The horizontal wind load « - 2228 X 25 - 55,700 lb. - 27.8 tons. 

The moment arm of iS is y = — X 180 ( — ^' . ,^ — I « 81 ft. The wind moment Af = 

3 \ 8.75 + 16 / 

495 2252 

81 X 27.8 « 2252 ft.-ton8. A = —-- « 6.2 tons per sq. it. /« - ± — r - 8.7 tons i)er sq. ft. 

80.9 257 

Maximum compression on leeward side A + /« « 6.2 + 8.7 = 14.9 tons sq. ft. Maximum ten- 
sion on windward side /i — /a « — 2.2 tons per sq. ft. 

1 

25 5 
Foundation. The length of base I - 25.5', Ai - P - 650 sq. ft. Si - -— - 2763. Weij^t 

6 

of foundation based on 1.9 tons per cu. yd. is 266 tons. The weight of the 4>^" lining is, 36 X 11 
X 0.063 » 25 tons. The total weight of column, lining and foimdation is, Wi - 495 + 25 + 266 » 
786 tons. 

The moment arm for R may be assumed the same as before or 81 ft., then M » 2252 ft.-tons. 

7fi6 2252 

Pi =■ — - =» 1.2 tons T)er sq. ft. Pj « ---- ■- 0.81 tons per sq. ft. 
650 2763 

Maximum soil pressure Pi + Pt ^ 1*2 + 0.81 » 2.01 tons per sq. ft. 

Standard Specifications for Perforated Radial-Brick Chinmeys.* 1. Scope. The work in- 
cluded under this contract is to consist of all labor and material necessary for the erection 
complete of one radial-brick chimney in accordance with this specification, which shall become 
a part of the contract. The proposal shall include all scaffolding, cartage, unloading of material 
and removal of rubbish necessary to leave the chimney in a first-class condition ready for 
operation. 

GENERAL CONDITIONS: (Items 2 to 6 inclusive.) 2. Delivery. The chinmey will be built 

at located on the 

railroad. 

Note. In order that bidders may correctly estimate freight, labor and insurance rates, it is quite 
desirable that this space be carefully filled in. 

Material may be unloaded on owner's siding which is within 

of the chimney site. 

Note. The wheeling or trucking distance from point of transportation delivery to site of chimney 
should be as carefully estimated as possible. Do not give distance as the crow flies. This item is 
important as it affects the contractor's estimate considerably. 

3. Space. Sufficient storage room for chimney contractor's materials will be provided 
adjacent to chimney as well as unobstructed access from transportation delivery to the site of 

* Prepared by the M. W. K^Uogg Co. 



CHIMNEYS FOR POWER fiOILESS 183 

Bhimnwy tor ddivcsy and removal of nuiiteriala and tools. At leaet one side of chimney will be 

kii Eiee and open by the ofntn tor hoiating and nwking qiace until the chimney is completed. 

4. Water. The owners will provide the chimney contractor with necessary water within fifty 



FiQ 13. Dbuon or Bahui^Bbice Ohiknzt. 
feet of the site of the chimney free of expense to the chinmey contractor. From this point the 
cbimntv cmtractor will make his own boee connections if required. 

5, Wcrkmatuhip and MalerMt, All workmanship and materials shall be first claas. 



Ig4 POWER PLANTS AND REFRIGERATION 

Tlie chimney contractor shall fumish a competent foreman under whose superviflion the 
chinm^ wiU be built. Chimney must be built in a thorou^, complete and workmanlike manner. 

6. Time of Completion. The. chimney contractor shall state in bid the guaranteed number 
of working days in which he wiU finish the chinmey after the receipt kA signed oontxact and ap- 
proved drawings. 

7. Foundation,* 

Note, Aa a general rule the foundation can be built cheaper by the owner than by the chimnegr 
oontraetor, for the reason that other concrete work may be going on or under contract, and the owner 
then gets the benefit of a unit price for larger volumes of concrete. It may also happen that the owner 
has men who can do this work at odd times; whereas, the chimney contractor must pick up men and 
must supervise the work with an expensive foreman. Since the chimnQr contractor is held reqx>n8ible 
for the design of the foundation, it would seem more satisfactoiy all around for the owner to build the 
foundation. 

Proper foundation will be built by the owner from plans and specifications to be fuiniahed 
by the chimney contractor, who will, upon completion, give in writing his i^)proval of the foun- 
dation as being sufficient to sustain the chimney and fulfiU the guarantee. 

Note, In case, however, it is desired to have chimney contractor build the foundation, the follow- 
ing may be used. 

The chimney contractor shall fumish a concrete f oundati<m of piopesr depth and spread to 

safely sustain the chimney. The foundation shall be not loaded to more than tons 

per square foot, which is the safe bearing value as determined for this work. 

Excavating shall be done by contractor for foundation. 

Note. The nature of the ground will be more thoroughly understood by the designer than cook! 
be ascertained by the chimney contractor, and for the purpose of getting bids on an even basis this 
method of limiting the bearing value of the soil would seem the fairest way of securing i>ids. In case 
it should be determined after excavation is made that the foundation should be larger or smaller, a 
corresponding increase or decrease of quantities at a reasonable rate could be made. 

The concrete shall be composed of cement, sand, stone or gravel in the proportion of one 
part cement to two and one-half parts sand and five parts of stone or graveL It shall be de- 
posited in the forms in layers not to exceed six inches in thickness and thoroughly rammed into 
place. Concrete shall be a wet mixture. 

8. Design. The design of the chimney shall conform to the following dimensians as shown 
on drawing attached. 

Height above top of foimdation .feet inches. 

Minimum internal diameter feet. inches. 

The wall of the column shall have one strai^t and true batter foom top to bottom. Tlie 
wall thicknesses and section lengths to be as shown on drawing. In case the contractor's standard 
wall thicknesses should not be exactly as shown, a variation of three per cent will be allowed 
in either direction. 

Note, As a rule chimneys built round for the entire height of radial brick are the cheaper (see Fig. 
13), but it ie sometimes advisable, however, to design the chimney with what is known as base and column 
construction (Fig. 14). Three considerations affect this design, namely, width of flue opening, a 
desire on the part of the designer to have the lower portion of the chimney match the building walla 
in color of brick, contour, or for other architectural reasons, or, if chimney is located advantageously 
to point where building brick are cheap. 

The most economical height of base is approximately one-sixth the height of chimney and unless 
the base is designed to match building courses or on account of the flue opening coming at an untuoal 
height, this rule should be followed: The dimensions should be made to the nearest 5-foot level above. 
For example: 100' chimney - 20' base, 126' « 25', 150' - 25' or 30', 175' - 30', 200' - 36', 
225' i- 40'. A point to be borne in mind also is that there should be at least 3 feet of base above top ol 
breeching entrance. 

*NoTB. — Concrete foundationa cost approximately $6.00 to $6.00 per cubic yard in place. 



CHlMNEyS FOR POWER BOILERS 




I. 14. Radial-Bkice Chfuney v 



Crw atfon A-i 
B OcTAOONAl, Base. 



1l fnqumtly occurs that due to limited headroom the flue openins must be wider th&n could otherwise 
be dcsifnedp A rule for determining the maximum width of Due opeuiu^ into chimney bsses is as 
follows: Multiply the width of the chimney at the botU>m by the following factors; 



186 POWER PLANTS AND REFRIGERATION 

For round chimney bases 33%. See Fig. 13. 

For octagonal chimney bases ^^2%. See Fig. 14. 

For square chimney bases 50%. 

It is advisable to keep the width of the flue as narrow as possible in order to TrminfAin the highest 
stability through the flue opening. A good rule to follow is to make the flue opening at least twice as hi^^ 
as it is wide. 

9. Base, 

Note, If chimney is to be buHt with base and column construction, use the following: 

The base of the chimney shall be built (here fill in shape of base) in shape feet 

high of the dimensions shown on drawing of straight, hard, well-bumed, well-shaped common 
building brick laid in full bed of cement lime mortar, as herein specified. 

Note. If round for the entire height, specify as follows: 

The chimney shall be built of perforated radial brick for the entire height, as hereinafter 
specified. 

10. Radial Brick, All radial brick shall be best quality, molded from refractory clay, sound 
ringing, hard, well burned, well shaped, of reasonably even color and free from checks; made 
to closely conform with the circular and radial lines of the shaft, and shall be weather and add 
proof. They shall have a water absorption of not less than five per cent nor more than twelve per 
cent of their dry weight after immersion for a period of twenty-four hours; and shall have a 
crushing strength of not lesi> than six thousand pounds per square inch. The total amount of 
perforations shall not exceed one-fifth of the cross-area of the brick. One cubic foot of radial 
brickwork shall wei^ not less than one hundred and twenty pounds. The outside faces of the 
brick shall be of regular size, so that the general appearance of the brickwork wiU be neat and 
uniform. 

11. Lining, 

Note, The height of the lining for chimneys ia found in the following manner: 

For ordinary boiler work where the temperature does not exceed 800^ Fahr. the lining should be 
approximately one-fifth the height of the chimney. 

For temperatures above 800^ and below 1200*^, the lining should be one-half the total height. 

For temperatures above 1200** and below 2000^, the lining should be full height of perforated radial 
brick. 

For temperatures above 2000^ in general, the lining should be of a special high-grade fire-biick cut 
to radius. 

For temperatures above 1200^ it is well, however, to take up with a chimney expert the exact 
method of meeting the conditions. 

The chinmey shall have an expansion lining built of perforated radial fire-brick four and one* 

fourth inches thick, feet hi^ from a point two feet below the 

bottom oi the flue opening. The lining prevents the flue gases from coming in contact with the 
solid masonry of which the shell is built, and shall be separated from same by an air space of not 
less than two inches. 

This lining shall be built after the chimney is finished and exceptional care must be taken 
to keep the air space clear and free of loose mortar and other dirt. 

Rack out the shell of the chimney approximately two inches above the lining to form a ledge 
for the purpose of diverting the falling soot when the chimney is in operation. 

12. Mortar, All brickwork shall be laid in cement lime mortar as hereinafter specified with 
courses level and with fiill mortar joints throughout. Face brickwork and backing to be laid 
up at the same time with joints of reasonably even thickness, not exceeding one-half inch. The 
mortar to be used in the chimney shall consist of one part Portland cement, two parts fresh-burnt 
lump lime mortar and five parts clean, sharp sand. The cement to be added to the sand and lime 
mortar as the mortar is required, and no mortar having taken an initial set is to be used- The 



CmMNBYS FOR POWER BOILERS 187 

cement must not be added until the lime is cool. The sand shall be clean and sharp, free from loam, 
vegetable matter and large pebbles. If necessary it must be both screened and washed. 

13. BontL All common bridcwork shall have every fourth course a header course. 

NaU, The above sentence may be omitted in case chimney is designed round for the entire height 
as per Fig. 13. 

Radial brickwork shall be bonded eveiy three oouraes. 

14. Breeching Opening, 

NoU, As a rule one opening into a chimney is sufficient. Sometimes it is desirable to locate the 
chimney in such a position that gases will be drawn in two opposite directions. It is then advisable 
to have two openings in the chimney. In case there are two openings a baffle wall from a point two 
feet bdow the bottom of the flue opening to five feet above the top of the higher opening should be 
provided. Baffle wall should be not less than four inches thick of refractory material and bonded to 
the Hning. When baffle is necessary see Fig. 14 for method of construction. 

One opening shall be provided in chimney. The opening to be lined on the reveals with 
refractory materiaL The masonry above the opening to be supported by heavy I beams set 
on steel plates with air spaces at each end for expansion. Under these I beams a flat masonry 
ardi shall be buHt to properly protect the beams from the effect of the gases. The flue opening 
shall be reinfctfoed laterally by heavy tie rods and plates over the top and at the bottom. 

Three-eighths by three-inch steel bands to be placed in the masonry above and below 
opening. 

The opening shall be wide by high, the bottom of which shall be 

apimmmately above foundation. 

Note, The area of the flue opening should at least be 7 per cent more than the area of the in- 
ternal top diameter of the chimney* but not more than 15 per cent. The width of opening should not 
exceed 33 per cent of the outside diameter of stack at base. 

15. Reinforcing Rings, The chinmey contractor sl^all place in the brickwork at every change 
in wall thickness steel bands three-eighths inch thick by three inches wide. 

If the contractor should furnish perforated radial brick having corrugated sides, these bands 
may be omitted. 

16. Head. The head of the chimney shall be neatly corbelled out and fitted with a heavy 
annnVM' retaining ring set in full bed of cement mortar. 

Note, If ornamental design at head of chimney is wanted, as shown on front cover, specify that 
same is to be worked in, using kiln-burned brick. 

17. CUanovi Doer. Provide and place in base of chinmey where directed by the owner a 
east-iron deanout door and frame properly hinged and fitted with latch. Said door to be approxi- 
matdy twenty-four inches wide by thirty-six inches high. 

18. Ladder, Build <m the interior of the chimney a ladder to consist of three-quarter inch 
galvanised iron rungs spaced approximately fifteen inches center to center and secively anchored 
to the masonry from top to bottom. These ladder irons to be in shape of a " U " with hooked ends. 

The ground plate shall be buried by the contractor for the foundation when it is buflt. 

19. Lightning Condiuior, The lightning conductor is to consist of 

Note, Two points are the minimnm for any diameter chimney up to five feet inside. Above 
five feet one point should be added for every two feet in diameter or fraction thereof. 

copper points three-fourths mch in diameta* by eight feet long with one and one-half inch platinum 
Ups. The points to be anchored to the top of the columli and extended from the bottom of the 
corbelling upward. The lower ends of the points to be connected by a loop of copper cable en- 
circling the chimney. From this loop there is to be one one-half inch seven strand No. 10 Stubbs' 
wire gage copper cable carried down the side of the chimney and connected to copper ground plate 
of the three-winged type as best for the proper distribution of charge. The points to be securely 
fastened to the top of the chimney and the cable to be anchored every seven feet in hei^^t with 
brass afiohors, eo designed that they will support the weight of the cable. 



188 



POWER PLANTS AND REFRIGERATION 



20. LeUering, 

Note. It frequently happena that an owner as an advertifling medium deaireB to haTO the 
or the initials of the company woriced into the chimney. This may be done at riishtly eztn 
The spedfioations should be as follows: 

Woric mto the column on (one or two) sides as directed the letters (here insert the desired 
legend) to be made in pennanently colored kiln-burnt brick. Letters to be true to sise and shape 
and to be in a true vertical line. 

21. Trimmings, 

Note. Due to architectural reasons on public buildings or in select residential districts. It is 
times desirable to have the chimney present a more ornamental appearance than is usual for 
factory work. Chimney shafts may be designed with either straight batter <»- with an entasis. The 
base and head portions may be decorated with either stone or terra-cotta courses. If other stone or 
terra-ootta work is to be done at the time chimney is to be built, it should be specified that these courses 
will be furnished by the building contractor to the chimney contractor. This is the most eoonomicsl 
way of handling this type of construction. Since the ordinary rigging employed by ctunmey builders 
is necessarily quite light, it should be borne in mind that no one piece of stone or teira ootta riiould 
weigh over 200 pounds. 

All necessary stone or terra cotta shown on drawing will be furnished without charge by the 
building contractor to the chimney contractor, who will set same. No one piece will weig)i over 
200 pounds. 

22. Insurance, The chimney contractor shall carry at his own expense during the entire 
period of construction liability insurance, insuring the men in his employ, and the public in 
general, in case of damage due to accidents. 

23. Ouarantee, The chinmey contractor shall guarantee the chimney for a period of five 
3rears from date of completion. The guarantee shall cover any defects that may arise within 
this period due to faulty design, construction, materials, weather, and the products of corn- 
bastion up to SOO"* F., 

Note, The guaranteed temperature should be dependent on the work to be performed (see note 
in Paragraph 11). 

and shall further guarantee to make good at his own expense all defects that may arise horn msy 
of the above conditions within the specified period. 

The chimney shall be designed for a wind velocity of not less than one hundred miles per hour. 

Cost of Radial-Brick Chimneys. OtMuxrdt, in his " Steam Power I^ant Engineering," gives 
the following costs of a well-known make of radial-brick chimn^:* 

TABLE 20 

COST OP RADIAI^BRICK CHIMNEYS 



Sob or Chimnby 

* 




SiZR OP Cbucnky 




Heislit 


DUuneter 


Coet 


Height 


Dimmeter 


Cost 




Feet 


Feet 




Feet 


Feet 








76 


4 


$1,360.00 


176 


8 


$7,060.00 






76 


6 


1.960.00 


176 


10 


7.926.00 






76 


8 


2.660.00 


176 


12 


8,960.00 






76 


10 


8,726.00 


176 


14 


9,726.00 






126 


6 


8,600.00 


200 


8 


9.260.00 






126 


8 


4^60.00 


200 


10 


10.600.00 






126 


10 


4,676.00 


200 


12 


11.100.00 






160 


12 


6,125.00 


200 


14 


12.600.00 






160 


8 


6,160.00 


260 


10 


16.600.00 






160 


10 


7,126.00 


250 


12 


18.260.00 






160 


12 


7.760.00 


260 


14 


21.600.00 






160 


14 


8.276.00 


250 


16 


24.260.00 





*The sctusi oQBt TBriee with the freight rste from the nearest point of manufftetvre. Prie«b «>Miaeive of foonda- 
tioa, in Buffalo. N. Y., 1916 were approximately 40% leei than eteted by UUe. 

Coet of concrete foundations fai place may be estimated at approximately $6.00 per ea. yd.; CKcaTstftOM^ %IJ0% 
perett.yd. 



CHIMNEYS FOR POWER BOILERS 



189 



REINFORCED CONCRETE CHIMNEYS 

Tlus type of dumney, althou^ of oomparatively recent dedgn, 10 finding much favor among 
en^neers and ita use is rapidly increasing. 

Concrete is well adi^>ted to resist compressive stresses, but quite inefficient in tension. In 
order to supply this deficiency in tensile strength, steel rods, bars, or structiual shapes are im* 
bedded in the concrete. 

This type of chimney occupies somewhat less space than a brick chimney on account of 
the thinneas of the walls at the base, and is much lighter in weight. It is a monolithic structure 
inasmuch as the foundation and stack are cast integral without joints. The construction of 
this tjrpe of diimn^ is quite rapid, being at the rate of about six feet per day. 

Ccmcrete chinmey shells are built both tapering and straight. The shell, if made straii^t, 
is ordinarily six inches in thickness, and if ti4)ered the shell at the top is made four inches in 
thickness and is increased in thickness one-quarter inch tar each five feet in height. See Fig. 16. 

Sufficient vertical sted reinforcement is provided to take care of the entire tensi<m on the 
windward side, resulting from the wind moment, the allowable stress in the steel being 16,000 
flb. per sq. in., giving an apparent factor of safety of about four. 

The maximum compression of the concrete resulting from the dead load and wind moment 
is ordinarily limited to approximately 350 lb. per sq. in. at the base of sheU. 

The following formula and example in the design kA a tillering reinfcnrced concrete chimney 
175 feet high by 9 feet diameter are given by the Oeneral Concrete Construdion Co.: 

Strain Sheet for Chiamey i75'-o'' His^ by g'-o" Diameter. 



Bdgiit of CUmMV Abov« Grade. . . 
D<|ith of FwhidartflB B«low Grade 

n d ght of FmmdMtkm to CcBtar 



IWO^ 


S'O" 


181' 0" 


24' Cr 


4' 6" 



DnoNnoNi 



Indde DiamtCer of Chimney at Top 

LarffMt Outiide Diameter 

Heiglit of lininiE 

Thidmem of Shell Tapen from lOH'' to .. . 
Aeeiamed Wind rreee ur e 60 U>. per Sq. Ft. . 



yO'* 

14' S" 

OO*©" 

6" 



JSymboU used 
Ac 
Ah 

AIt-AU 

D 

Dw 

d 

A 

d, 

Wf 

V 

We 

Ws 

Wl 

Wi 

Pd 

P 

L 

M 

Fw 

Fm 



in Strain Sheet: 

>- area of concrete in section in sq. in. 

s area of cross-section of 4yi" lining. 

>- areas tor 9", 13", etc., lining. 

« outside diameter of shell at section. 

« outside diameter of sheU at bottom of portion exposed to wind. 

as inside diameter o( shell at sectimi. 

» outside diameter of shell at top of chimney. 

« inside diameter of shell at top of chimney. 

» weii^t of foundation. 

>B cubic feet of space occupied by chimney below grade. 

s wei^^t of earth on foundation. 

« wei^t of shell above section. 

s weight <^ lining at 120 lb. per cu. ft. 

» total dead load <m soiL 

>- eart^ pressure from wei|^ 

s total wind pressure above seotioo. 

» lever arm of wind pressure. 

» wind moment at section. 

"B earth pressure from wind. 

» maximum earth pressure from wind and weight. 



fill 

mil 



RieVllanr Sttl 12 '^^tt I" btOi lUneUBn 24 Cob 
IXatoatI StttI It .fUtt balt> ■ijw on tCtatui. 



CHIMNEYS FOR POWER BOILERS 191 

R « radius of neutral core. 
M8 « moment of stability from weight. 
Mh SB bending moment at section. 

/ » moment of inertia of section. 
Re B extreme fiber distance at section. 
Fc >e extreme fiber stress in concrete from wind. 
Fct « stress in concrete from weight. 
Fct >" total compression in concrete from wind and weight. 

t » toision allowed in steel. 
Ra a radius of steel circle. 
S » percentage o( steel in section. 
Fs n compression in concrete, taken by steeL 
Fcm » maximum compression in concrete. 
H s height of shell above section. 
h *B height of shell exposed to wind. 
As s area of steel in section. 

Section at Bam cf FaundaUan. 

L— W/- <^X2-f ^ '*"g^^'^ X2.6) 160- 848.090 Uk 

n-— V -2.125.88+289.77- 2J66* Cu. PL 

m.— ir«- t(2?X6)- V)J100- 109,0i0 Lb. 

IV.— 1P«- |[I>» + (DA)+A«l-[d»+(ddi) + d»«l| 0.2618H160- 827,861 Lb. 

v.— m - 79.488 Lb. 

VL-r-ITI- ITZ + ITc + ira+iri- 1.164.469 Lb. 

Vn^-Pd--^- 1.971 Lb.Sq.Ft. 

VnL-P -5!»+^x*-y-- 6^.688 Lb. 

-(^?^Xt)+«- «"^«- 

m PL-^ , 4,688,787 Ft^Lb. 

XL— Pw -^- 2.018 Lb. Sq. Ft. 

XIL-Pm- Pd+Pw « 8,984 Lb. Sq. Ft 

Section at Top cf Foundation. 

&-ir*- |[D> + (DDi) + Dil ~Id> + (<Ui)-f di>]j 0.26180160- 627,861 Lb. 

n-^ -f[l+(^)*]- 8.147 FL 

m^-rift- /nra - 1.975347 Ft.-Lb. 

rr^p .5!»+^x*^- 62.588 Lb. 

^-'^-(^;^xt)+^«- «-^^ 

VL—M '^PL ^ 4.402.141 Ft-Lb. 

Vll^-Jtfb-Jf-Jfa - 2.426.294 Ft^Lb 29.116,628 In^IiK 

Vm^-/ - 0.049 (D« -4f<) valuM in inehM 17.276.635 

IXr-Re- 85mn. 

X^-fe . i^£S. 144.1 Lb. Sq. In. 

XL— Ae - 5.863 Sq. In. 

XIL— Pc!,-— - 1173 Lb. Sq. In. 

ZnL— P€^ - fc +2^Cfe - 878.7 Lb. Sq. In 

XIV.— I - 16.000 Lb. Sq. In. 

XV.-«f - 82HIn. 

ZVI^A* - ^^ - 44.1144 Sq. In ^^g - 78.4 BAni 



192 POWER PLANTS AND REFRIGERATION 

Section at Top of Foundation — (Ccoitinued) : 
XVII.— s -l^Jlf- &a 

xvm.— T. - ^ - j^ — - su Lb. sq. In. 

XIX.— Fm- ^e, - ft ML4 Lb. Bq. In. 

Al - llMSq-rt. D - U.ZSFt. d - II.tSFt. 

AI,- 8q.FL Oi - 9.SS Ft. *- B.OOFI. 

Ail- Sq.Ft. Dit - I4.3IFt. S-.... ITCM Ft. 

k - 1T6.00R. 

SpodflcatloDS ias a Coner«t« C]iinuie7. 

Dimenauma febt inchzs 

Height of chimnejr above grade 

Depth of foundation below grade 

Total height 

Size of foundation 

Height of foundation in eent«r 

Inaide diameter of chimney at top 

I^^eat outside diameter 

Height of lining 

HiidmeM of riiell, tapering from inches to 

Exemaiion. Purchaaer shall do all excavating, protect embankment, keep foundatiiw pit 
free &om mter and do axty piling that may be required. 



Ddwery and Time. Materials and tools will be shipped in days from receipt 

of notice; unloading to be done by purchaser on arrival. About days wiU be re- 
quired to complete t^ work. 



CfflMNEYS FOR POWER BOILERS 193 

Water and Space. Purchaser Bh;^ furnish a supply of clean water within SO feet of the base 
of the (duimiey for the prosecution of the work, also dry-etorage room for cement and toots, and 
ample space for other materials. At least one side of the chimney shall be left free tor hcHsting 
material and mixii^ concrete until work is 
ciMiipleted. ~ 

MateriaU and WorkmaTuhip. All the mate- 
rials will be the best of their respective kinds, 
the Portland Cement will be of a standard brand 
f ulfilling the specifications of the Amaieart 
Sodely jar Tt^ng MaieriaU. The work will be 
done in a firat-clasB woi^manlike manner under 
the BupervisioD of an experienced foreman. The 
concrete will be thoroughly mixed and tamped 
in the forms and around the steel to secure the 
beet possible bond. The chimn^ will be built 
with our patented all-steel forms, insuring a 
snootfa and uniform surface on the concrete, 
which BtUsT completion will be given a coat of 
cemoit wash. 

Beinfarcemmt. The foundation will be rein- 
tonxA with two nets of three-quarter inch square 
twisted steel; the lower net placed diagonally and 
steel spaced twelve inch centers; the upper net 
placed parallel to sides, steel spaced twenty-four 
inch centers. The vertical reinforcement in the 
chimney will consist of thre&^uarter-inch square 
twisted steel; sufficient bars will be used to 
■bscnb all tenmon without stressing it beyond 
16,000 pounds per square inch. Rods will be uni- 
fon&ly spaced, and placed 3 inches from the 
outer surface of the concrete. Joints will lap 30 
incbes. Tlie vertical rods will be embedded in 
the foundation and bent under foundation steel ~ 
for anchorage. Hie horizontal reinforcement will 
be a steel net consisting oi on».quarteHnch 
longitudinal rods spaced four inch centers, trian- 
gularly laced, the ends lq>ping 6 inches. This 
net will be placed around and wir«d at intervals 
to volical steeL 

Coitavle. The concrete in the foundation 
win be mixed in the proportion of one part of 
Portland cement, three parts clean sand, and six 
parts crushed stone or graveL The concrete in 
the chimney will be a "wet mixture" of one part Portland i iiiufTit, two and one-half parts clean 
sand, and three parts of one inch crushed stone or gravel. 

AUackmentt and Idning. The chimney will be provided with openii^ for flue connection 
and a CBstr^ron dean-out dow. The lining will consist of a good grade of hard-bumed brick, 
covered with a ocxicrete cap, and separated from the concrete shell by an insulating air space. 

Detign Wild OuiranUe. The foundation will be of such size that the resultant of forces will 
fall within the middle third, and the mftinT""" compression from live and dead load will not 
exceed the Baf&4>earing value of the soil. The shell at the base of shaft will be of such thickness 
that the mBirimum compression on concrete will not exoeed 350 pounds per square inch. At tho 



Pro. 17. 



IM 



POWER PLANTS AND REFRIGERATION 



emcke Gpening the thickness of shell will be increased about 30 per cent on each side and e3cteiiding 
five feet above and below, and additional reinforcement provided. The chimney wiU be designed 
to withstand a wind pressure due to a wind having'a velocity of 100 miles an hour and chimney 
gases not exceeding 1000 degrees F. For a period of five years after completion we will repair 
free of charge any defects arising from faulty design, defective materials, or workmanship. 

Cost of Reinforced Concrete Chmmeys. The following data compiled by H, A. Strattst are 
given to convey an idea of the selling prices of modem concrete chinmeys. 

Table 21 represents a series of chinmeys on the basis of which estimates of the probable 
cost <A a chimney oi this type may be made. Table 22 is a rec(»d of actual installations of this 
type of chinmey. 

It is difficult to give a list of prices that will be general, because the cost of materials and 
labcHT vary locally and such chimne3rs are invariably manufactured and erected at the site. 

The prices given apply in general in the Central and Eastern U. S. A., but a 10 per cent 
increase in the prices given should cover practically any case, however remote, if located on a 
Bteam railroad with throu^ connections. 

Construction Data. Foundalum, (a) Concrete consists of one part Portland cement, three 
parts sand, five parts crushed stone or gravel (not over 2" sice). 

Q>) Reinforcement consists of steel bars. - 

(c) This foundation is included in the prices given <hi this card. 

(d) Excavation for foundation is not included in price, nor piling if required. 

Shaft, (a) Concrete consists of one part Portland cement, two and one-half parts sand, four 
parts crushed stone or gravel (not over 1" size). 

(6) Reinforcement consists of vertical steel bars and horizontal steel rings. The latter take 
up temperature stresses and also shearing stresses caused by wind pressure. 

Openings are provided for flue connection and clean-out door. 

Designed for wind pressure of 100 miles per hour; and chimney gases up to 1200^ F. 

Designed for total load on soil of i^proximate 2 tons per sq. ft. 

For underground, flues (A) and (D) are increased. Use next higher price to allow for this. 

TABLE 21 

REINFORCED CONCRETE CHIMNEYS 

APPBOXIMATB PRICaBB 



Heiglit 
AboTS 
Gnuli 


Inride 
Diameter 


Depth 
Below 
Grade 


Hfliriit 

Double 

SheU 


Height 
Sinfle 

shSa 


Total 
Height 
A-B-C 


Otttdde 
Diameter 


Width o( 

Square 
Foundation 


Approxiauit« 

ScUinc 

Price 


H 


G 


A 


B 


C 


D 


B 


F 


1 


100 
126 
126 
160 
160 
160 
176 
176 
176 
200 
200 
200 
226 
226 
226 
260 
260 
260 


4' 

6 

6 

6 

6 

7 

8 

8 

9 

10 

10 

11 

12 

12 

18 

14 

14 

16 

16 


6* 
6 
6 
6 

I 

8 
7 
7 
7 
7 
7 
7 
8 
8 
8 
8 
8 
8 


88' 

88 

42 

42 

48 

48 

48 

67 

67 

67 

66 

66 

66 

69 

69 

69 

81 

81 

81 


• 67' 
67 
88 
88 
102 
102 
102 
118 
118 
118 
184 
184 
184 
166 
166 
166 
169 
169 
1^9 


106' 

106 

180 

180 

166 

166 

166 

182 

182 

182 

207 

207 

207 

288 

288 

288 

268 

268 

268 


6' 4" 
7 4 

7 4 

8 4 

8 4 

9 4 
10 4 

10 6 

11 6 

12 6 
12 6 
18 6 
14 6 
14 8 
16 8 
16 8 

16 8 

17 8 

18 8 


12' 

12 

16 

16 

18 

18 

19 

22 

22 

28 

25 

26 

26 

29 

29 

80 

82 

88 

84 


$2,000 

2.800 

2.800 

8.200 

4.000 

4.600 

6.000 

6.000 

6.900 

7.800 

9.000 

9.900 

10.800 

12.200 

18.200 

14.200 

16,800 

16,900 

18,000 



CHIMNEYS FOR POWER BOILERS 



195 



TABLE 22 

REINFORCED CONCRETE CHIMNEY ACTUAL INSTALLATIONS 



Hdi^ifc StXMTC 

Depth found balow crade 

Width aquMre pmrt of foundation 

Btt^^ mqampmrt o( foundation 

tan oiiMda diamotcr 

ipow9T installation 

power maadmum dwInKOTCfload 

l o ca tion (Stnt«) 

Actual priee wcc ttwl comiJcte 



D 


160^ 


H 


164' 


m 


160* 


A 


6* 


F 


22* 


• ■ 


8' 6^ 


G 


ye" 


B 


10*10" 


• • 


1.200 


• ■ 


1.800 




Ohio 


• . 


$4,800 



171' 
166' 
160^ 

e' 

20* 
8'«" 

10'4" 
1.200 
1.800 
Ind. 

$6,800 



187' 

180' 

176* 

7' 

28' 

8' 6" 

10'6" 

18' 10" 

2,400 

8.600 

Ind. 

$7,600 



216' 


266' 


210' 


260* 


200' 


250' 


er 


6' 


82' 


87' 


4' 6" 


4' 6" 


16* 


15' 


17' 8" 


17' 8" 


8.000 


8.000 


12.000 
CaL 


12,000 
CaL 


$14,760 


$17,600 



NOTB.— ' 

topis'. 



larcaatehfaniMy of tfaittjpafai the world Btaada at Butter Mont. It li afiO* fai hMit, iniide diameter 
1906. 



CHAPTER Vra 



MECHAinCAL DRAFT 



General Conditioiis. The rate of driving a boiler plant is dependent upon the intensity 
of draft available, which, with a chimney, is limited by the height and temperature of the flue 
gases and is, furthermore, somewhat susceptible to atmospheric conditions. 




§ i S § S 

«« ^ «» 00 § 



o o o o 

o o o o 

> so Qo Q 

-^ -^ -^ CM 



Pounds of Co*/ per Hour 



Fig. 1. Belattvu Heigbtb. Diambters and Oostb of OnncNETS for Tatlor and 

Natural Draft Furnaces. 



Artificial or mechanical draft is not subjected to any of the above-mentioned limitations and 
has, briefly, the following advantages over natural draft: 

Independent of climatic conditions. 

More readily controlled to meet the varying demimds of load. 

Permits the use of cheap low-grade coal which requires an intensity of draft beyond the 
hei^t of reasonable chinmey construction. 

When the height of chimney is not limited by local ordinances, fan draft equipment for 
medium and large size plants is ordinarily considerably cheaper in first cost than the equivalent 
chimney installation. 

196 



MECHANICAL DRAFT 197 

In Kuural, where local ordioancea require chimneys of considerable height, artificial draft 
eqoipment is not ordinfuily installed except in large central-etation work to meet emergency 
peak kisda or when a forced draft type of sbiker is installed. Forced draft is frequently installed 
in old plants in order to force the boilers beyond tbe capacity of the chimney installed to save 
the expense of additional boilers. 

A dbimney, when once erected, costs nothing for operation, while the operation of any type of 



Pio, 2. Section THRoran Boiur, BsowiNa How Bloweb la Imbtaujd 

mechanical draft apparatus requires the use of steam which amounts from 1^4 to 6 per cent of the 
total Bteam generated, depending upon the size of plant and method of driving the fans. 

The curve, Fig. 1, prepared by the manufactureis of the Taylor stoker, may be used in 
appronmating the saving in cost of chinmey when forced draft is used. The friction losses are 
anumed at 200 per cent boiler rating. A ctmstant height of 100 feet for the chimney is assumed 
for tbe forced draft equipment. 

Classiflcation. Artificial draft equipment is clasmfied as either Forced or Induced. With 
farced draft a pressure is created in the ash pit by means of a fan or steam jet, the air being forced 
thnw^ the fuel bed. 

With induced draft, a partial vacuum is created in the furnace by either of the two methods 
moitioned, the air being drawn through the fuel bed, producing tbe same effect as forced draft. 

Steam jet blowers are not used except in small plants and principally for the purpose of 
qoickly raising steam m small portable boilers. 

Pan Draft. Fan draft equipment is invariably installed in medium and large size plants 
in which artificial draft is employed. Tight boiler settii^ should be tbe rule in any plant for 
the most economical operation and are imperative with fan draft equipment. 

The main difference between forced and induced fan draft ties in the difference in volume to 
be handled by the fan. The weight of gas is the same in either case, the volume, however, with 
forced draft is based on the temperature of the air in theboiler room, say 70° F., while the volume 
of gases to be handled by an induced draft fan is based on a temperature of approximately 550° F. , 
the ratio of volumes being approximately 1 is to 2. The ratio between the fan speeds necessary 



198 



POWER PLANTS AND REFRIGERATION 



to prcxluce the same pressure (see the Chapt«r on "Hot Blast Heating," Volume I, Table 27) k 
1.38 times the speed required (or air at 70° F. 

A two or more (an equipment ia always advilsble in medium and large siie plants which 
must operate oontinuoualy or irtkenerer a shutdown of the draft system would praveoit cairyiug 
the load. 

FORCED DRAFT 

Forced Draft for Small PUnta. A type of f<wced draft equipmrait, particularly Bd:^>ted tor 
email plants, is a oombinaticm of steam turbine direct-connected to a propeller type fan installed 




Pia. 3, SBc?noN TaaoDoa Stcstzvant UNDiaaaATa Bloweb. 



in the ude-wall setting beneath the grate. A typical example of undergrate blower ia ■ 
by Figs. 2, 3 and 4. Installation data ore given by Tables t and 2. 





""SSZT" 


A 


B 


c 


D 


E 


r 


G 


' 


/ 


K 


I, 


tf 


S 






Stwn 


Ea.- 




i 


j 


f 


M 


r,H 


It'll 


17 H 

h 


S3H 

83 !i 


1 


1 


1 


ii 


i 


i 


B 


5g 


i 



Example. AsBume that a plant burning 9S0 pounds of coal per hour requires 25 per cent more 
power. Approiimatdy 25 per cent more coal must be burned or 080 X 1-25 ~ 1225 lb. per hour. 

If the grate surfaoe is 35 sq. ft. the rate of combuition R — 1225/35 — 36 lb. per sq. fU per hour. 
ReferriDg to the curves, Fi|[. 3, Chapter IV, the draft requirementa for burning anthracite pea coal at 
this rata is 1.26 in. water static pressure. The diaft furnished by the ohiDmey may be neglected and 
IB taking oare of the prteeure less through the boiler, breeohing, and ohimner. 



MECHANICAL DRAFT 



199 



TABLE 2 

AND MAXI MUM VOLUME OF AIR DELIVERED BY VARIOUS SIZES OF 
STURTEVANT TURBO-UNDERGRATE BLOWERS 



IV 



of W«Ur 



H 



H 



iH 



IH 



IK 



2,725 

8,820 
4,870 



2,725 
8,276 
8320 
4,870 



2,725 
8,275 
8320 
4370 



8,000 
8,660 
8320 
4.080 
4370 



8375 

8360. 

8320 

4.080 

4370 



8350 
8320 
4.080 
4370 



8,820 
4.080 
4370 



18" 



H 



H 



IK 



IH 



2370 
2320 
8370 
8310 

2,470 
2,920 
8370 
8310 
4370 

2,470 
2,920 
8370 
8310 
4370 

2370 
2320 
8370 
8310 
4370 

2320 
8370 
8310 
4370 



Co. Ft. 

of Air 
Supplkd 
«r MbniU 



1.660 
2,075 
2,625 
2,920 



1386 
1386 
2325 
2316 



890 
1,780 
2326 
2,746 



980 
1,780 
2300 
2,440 
2,660 



990 
1376 
1,960 
2385 
2,666 



1.070 
1,690 
2,080 
2320 



1320 
1.700 
2300 



2326 
8320 
8,960 
4,660 

2,620 
8.190 
8.780 
4.480 
6.060 

2.066 
8.020 
8.710 
4380 
4360 

1365 
2,520 
8.560 
4.180 
4,800 

2.155 
8380 
4.120 
4.780 



Fwn 



18' 



22^ 



26 



[" 



Stade 



inlnehM 
ofWator 



IH 



2 



24 



H 



H 



IK 



IK 



IK 



2K 



8 



K 



R^PsoK* 



2.920 
8370 
8310 
4370 



8370 
8.810 
4370 



8370 
8,810 
4370 



1306 
2390 
2.670 
8.050 



1305 
2390 
2.670 
8,060 
8385 
3,816 



1,905 
2390 
2.670 
8.050 
8.436 
8.815 



2390 
2,670 
8.060 
8335 
8315 



2390 
2.670 
8.060 
8.485 
8,816 

2,670 
8.050 
8.485 
8.816 

2.670 
8.060 
8.435 
8.815 

8.050 
8.435 
8.815 

8.060 
8.485 
8.816 

1.560 
1.710 
2,020 
2340 
2.640 



Cu. Ft. 

of Air 
SoDDlied 
er Minal 



at« 



1365 
8.080 
8.940 
4.620 



2.4S0 
8.760 
4.600 



1.650 
2360 
4.180 



8380 
4360 
6.180 
5380 



2,940 
4.070 
4370 
6,770 
6,670 
7,626 



1380 
8.660 
4.770 
6.680 
6326 
7.380 



2.880 
4370 
6360 
6360 
7300 



2.080 
4.025 
5360 
6326 
7.080 

8365 
4.970 
6.980 
6.960 

2.600 
4.610 
6330 
6.760 

8.000 
5.120 
6.600 

2300 
8.940 
6,880 

6.000 
6.900 
6.960 
8.800 
9.500 



SiM 

Fan 



26 



// 



Static 



inlndiM 
ofWator 



K 



IK 



IK 



IW 



2 



2K 



8 



8H 



R.P.aKa 



1,660 
1.870 
2.180 
2.490 
2.800 
8,110 



1,660 
1370 
2,180 
2,490 
2.800 
8,110 



1,710 
1370 
2,180 
2,490 
2,800 
8,100 



1.870 
2.180 
2.490 
2.800 
8,110 



2,020 
2.180 
2.490 
2.800 
8.110 



2,180 
2340 
2,490 
2300 
8,110 

2340 
2.490 
2.640 
2.800 
2360 
3.110 

2390 
2,640 
2.800 
2.960 
8.110 

2,490 
2.640 
2.800 
2.960 
3.110 

2,640 
2.800 
2.960 
8.110 



Cii.Ft. 
or Air 
Sap^ied 
MtaAxte 



4360 
6.060 
7350 
8.660 
9.960 
11300 



2,720 
6350 
7,100 
8390 
9,650 
10,900 



2380 
4380 
6,620 
8,120 
9,440 
10,700 



8,020 
6,000 
7340 
9,170 
10,600 



8300 
4.860 
7,410 
8,940 
10300 



8.720 
5300 
6.720 
8.690 
10,050 

3.460 
4.460 
6300 
7.630 
8.670 
9,700 

8.420 
4.470 
5370 
7320 
7,760 

2.800 
8.580 
4360 
5.480 
7,110 

2,910 
8.530 
4310 
6.440 



200 POWER PLANTS AND REFEUGERATION 

The weight of ur to be furnished by the fan m&y be BBSumed aa 20 lb. per lb. o/ co«J as 
(See tbe Chapter on "Fuels and Combustion.") 

The Bolunw of tit, meaaured at 70° P., to be handled l^ the fan per minute Ib thenfore: 
1225 X 20 , 



60 X 0.075 
Befenins to TaUe 2 a 22" mn fao operaliiig at about 3200 r.p.m. would be obowD. 



Flo. 4. DncsNBioN DRAwnra or Si-uui'iivant Vmnaa^Ara BiAwsa. (Table i.) 

Fomd Draft for Large Plants. When several boilers are to be served with forced draft 
in a new plant, a duct syBtem is inatallod beneath the boiler-room floor connected with one (w 
more fans, the outlets being located in front of the bridge wall, as indicated by Figs. S and 6, and 
controlled by a damper for hand^red installations. 

The duct system may either be constructed of brick, tile, concrete, ot galvanited sheet sUeL 



Large radius curves should be Knployed in all casea and aquai:&«<HTieied turns avoided to prevent 
exceeeive friction. 

Ducte are designed for air velocities of 2000 to 3000 ft. per minute. The pressure loss may 
be estimated, for a given layout, by the data given in the Chapter on "Hot Hast Heating," 
Volume I, or Equation (2), Chapter Vil, using a coefficient at friction / = O.OOU. 

Typical arrangements of forced draft equipment are shown by Figs. 7 to IOl 



DOublX-FAM ABHANanCMNT VOB A SWaiiB LlNB OF Bouxu. 




mm 



Fni. 9. Two-Pah ABBANaEUMr KB Docbui 



202 



POWER PLANTS AND REFRIGERATION 



Fig. 2, in the Chapter on " Mechanical Stokers/' shows a forced draft installation in con- 
junction with automatic stokers in which the speed of the fan is controlled and governed by the 
demand for steam. This may be accomplished in any fan draft installation, where a steam en- 




Fio. 10. 



gine or turbine is used for drivingy by the use of a pressure regulator controlling the steam 
supply to the engine. 

Size of Fan and Power Required for Forced Draft The following data are used by (me 
prominent fan manufacturer in determining the size of fan and power required for driving. 



MECHANICAL DRAFT 

Air Bequired per MinuU: 

2S ou. ft. at 70° per boiler horBepover with chain graUe. 
21 cu. ft. at 70° per boiler horsepower with ordinary grsUe. 
18 cu. ft. at 70° per boikr horeepower with underfeed atokera. 



Fio. II. PxarosicAMcB Cmtraa AmcaicAK Btbil Pi.it> Fanb. 

Preuure lUquirtd. Ordinary grates — 1^" water static preeeure with allowance of sufficient 
nr to speed up to \%" s.p. 



204 ^ POWER PLANTS AND REFRIGERATION 

Stoken — 2^" w&ter aUtie preoBure, Dot including duct friction. 

Where fan blows direct^ into aah int without ducts lH" s.p. wiU not be exceeded with 
oidinaiy rates of combustioii. 



Fio. 13, PiRroBUANCE Cuavse Avehican Sibocco Famb. 

Air HaruUtd. 253.5 cu. ft. at 70" F, per lb. ot coal burned. Thia amount oorrespooda to 
approximately 19 lb. ur per lb. of coal burned. (253.5 X 0.075 - 19.) 



MECHANICAL DRAFT 206 

Fan Data. See the Chapter on "Hot Blast Heating," Volume I, and also Figs. 11 and 12. 
Engine Data, See the Chapter on "Steam Engines." 

Example. Required the sise of fan and fan engine and the amount of power necessary for a forced 
draft installatton to provide for the following equipment: 8-150 b.hp. units steam pressure 150-lb. 
gage. Total rtUed capacity 1200 b.hp. Assume that the coal used has a calorific value of 12,000 
fi.tu. per lb. and the overall efficiency of the boiler, grate and furnace is 65 per cent. iWperature of 
feed water 170** F. 

Heat required to evaporate one lb. of water for the assumed conditions is: 1196 — 138 » 1058 
B.t.u. 1 boiler horsepower -> 33,524 B.t.u. 

33,524 
Tlien ^^^ ™ 31.7 lb. water to be evaporated per b.hp.-hour. 

_ 12,000 X 0.65 

Water evaporated per lb. of coal Is: — — ■■ 7.4 lb. 

1058 

Coal required per b.hp. « 31.7/7.4 b 4.3 lb. per hour. 

Total coal required per hour » 1200 X 4.3 « 5160 lb. at normal rating of boflers. 
If a 50 per cent overload is to be guaranteed then the coal required per hour will be approximately 
5160 X 1.5 a 7740 lb., and on a basis of 19 lb. air per lb. coal the volume of air to be supplied by the 

. . 7740 X 253.5 ^„^^ ,^ 

fan is: -■ 32,702 cu. ft. per mm. 

Aasuming ordinary hand-fired grates are tised, a static pressure of 1^" water at the fan will ordi- 
narily be sufficient. 

Referring to Fig. 12 it will be found that a No. 10 fan, operating at 220 r.p.m., will deliver 
32,500 cu. ft. per min. and requires 12 brake horsepower for the assumed pressure. This will require 
a 7'^ X 7" engine based on 150 lb. gage pressure (see the Chapter on "Steam Engines.") The water 
rate for this sise automatic engine will run about 35 lb. per i.hp.-hr. 

Awwiming a mechanical efficiency of 90 per cent for the engine and 1^" s.p. for the fan, the steam 

12 X35 
oonsumption of the forced draft equipment will be: — -— - — ■■ 466 lb. per hour. 

vF.jlU 

466 

This ifl -— -— - X 100 or 0.82 per cent of the total steam generated. 

1200 X 1.5 X 31.7 *^ 

It may be safely assumed that the steam consumption will be increased to approximately 1 H percent 
after equipment has been in operation for some little time. 

Some engineers prefer to install a larger engine than is actually required in order that fan 
may be opiated at full speed with a low steam pressure. 

The main objection to this is that with a fixed cut-off engine used in this manner the efficiency 
is lowered, due to wire-drawing of the steam at the governor throttle when operating under full 
steam pressure. 

A Diore refined method of calculation is to base the static pressure rating for the fan on the 
sum of the estimated losses through the duct system, fuel bed, boiler, breeching, and chinmey. 
The pressure loss in the duct ssrstem may be estimated from the data given in the Chapter on 
" Hot Blast Heating," Volume I. The loss through the fuel bed and boiler is given in the 
Chapter on "Pow» Boilers," and the loss in the breeching and chimney in the Chapter on 
"Qumneys for Power Boilers" of this volume. 

INDUCED DRAFT 

Ordinaiy Systems of Induced Draft A typical induced draft fan equipment is shown by 
Kg. 13. 

The principal advantage of induced draft over forced draft lies in the fact that it is not 
necessary to shut off the draft when cleaning fires with hand-fired boilers and some t3i>e8 of 
mechanical stokers. The following data may be used in calculating an induced draft installation. 



206 POWER PLANTS AND REFEUGERATIDN 

TempertUuit ^ Gaaet. 550° F. without econraniierB, and 360° F. with eamcMiiiiets. 

VoIum« qf Qa»ti based on 19 lb. air per lb. of coal, the volume <rf gaaee will be: 388 eu. ft. 
at 350° F. and 482 cu. ft. at 650° F. per lb. coal buraed. 

iSudton Reqaired. For rated capacity, 1" water static piewure; for 2G per cent overload, 
IJ^" water stAtio preeaure; and for GO per cent owload, 1 %" water static pRasure. 



Fia. 13. TmCUi InUVCED DBjUT lNBTAU.lTtON. 

rer necensuy for tlie boiler plant in the preeadtng 

The volume of gaaea getierHited per minute u: — 61,920 cu. (t. The atktic pTeseura 

rating of tan lo be IJi" wbIbt. 

For a conBtant preanue the ratio between the volume, speed and power required for a l«mperataie 

of 660° F. and the temperature at which the fan ia rated, or 70° F., ie:. I . 1.38. Thenfore 

\ 460 + 70 

a fan ia choeen having a capacity of 61,920/1.38 or 44,870 cu. ft. per min. at 70° F. and IK" B-P' 

Referrins lo the fan performaiice curves. Fig, 12, we (ind that a No. 10 Simoeo will deliver this 

amount of air against 15i" B.p. when running at 240 r.p.m., and requiring 20 brake horaepower. 

The speed and horaepowcr for 650° F. wiU bcSSO X 1.38 or 345 r.p.m. and 20X1.38 or 27.6 d.hp. 

If fan is to be motor-driven, a 30 lo 35 hp. motor would be used. 

The Prat System of ladnced Draft. A type of induced draft popular in Europe and known 
as the Prat i^Btem is shown by Fig. 14. The essential features are, the use of a double tapered 



MECHANICAL DRAFT 207. 

durnney, and a (an which does not t&ke all the ^laea from the boiler, but which may be arranged 
to cuise a draft on the inspirator principle, using either outside air or a part of the flue gues aa 
maybe demied. 

In the cotter of the chimney, just below the narrowest part, is fitt«d a noule connected to 
the fan, the bloat through this noule causing a suction in the lower part of the chimney. The 
object ot the t^>er is to give increased section for the outgoing gaaee, thus decreasing their 
speed and reducing the prtaoire at which the gases are discharged into the atmosphere. 



Tio. 14. Pkat are^ni or Inducw) Dratt. 
A. Noole for Pnt &7*Um. without dllhuer. B. Prmt Synem noule wlUt dUTuaer and aanolar 
dunpv. C. C row ■ectton of power plant, ■howhig "outot circuit" and "In circuit" metbod of io- 
■lallfiiS the Prat ByMem. 

In the "out-of-circuit" syBt«m the fan draws its air from the outer stmosphere, and in the 
" in-dreuit " system the fan is placed as a shunt to the flue drawing in and discharging into 
the narrowest pc^tion of the stock a portion of the flue gases. 

With the "in-circuit" the fan must handle the hot gases. In practice, the "out-of-circuit" 
method is used mainly for small installatione where there would be httle saving in power by the 
use of the "in-circuit" which is more commonly employed in la^e plants. 

By the "in-drcuit" method and the use of on inspirator, the fan employed is of relatively 
small capacity, being oaly about one-fifth the siie necessary if the fan handled the whole volume 
of gases. 



CHAPTER rX 

FEED WATER HEATERS AND FEED WATER PURIFICATIOn 

FEED WATER HEATERS 

The primftry purpoae of a feed water heater is to utilise part of the exhaust steam from so 
engine or turbine to raise the temperature of the feed water and ther^y return a p<Hti<Hi of the 
heat of the odiauBt, that mi^t otherwise be wasted, to the boiler. A saving of 10 to 12% in 
the fud is readily attained by the addition of a feed water heftt«r in a plant in which oold feed 



Pra. 1. Opbn Ttfb Hk&tbb> 

water ia used. The action of a feed water heater is simiUr to that of a condenser in that the 
Ut«nt heat ^ven up by the exhaust steam, which ia condensed, is used to raise the temperature 
of the water circulat«d through the apparatus. ' . 

^N feed wat«r heater serres another useful purpose in as much as it pKvents the feeding 
206 



FEED WATEK HEATERS AND FEED WATER PURIFICATION 209 

</ cold water into a. boilec and ther^y setting up a atieas due to unequal expansioti of the tubes 
>nd td&tes. 

Oasaificatioii of Pe«d Water HMten. Hie usual daaaification of heateta using exhaust 
aleam is nwde ftoomtling to the method ot the heat teansfer of the heat in the steam to the feed 
water and are aoeiMtliiiely known aa either the " Open " or " dosed " typee. 

Open Heaters (Fig. 1). In this claas belong all heaters in which the exhaust ateam nungleo 
directly with the feed water and whatever amount of steam condensed ia returned to the boiler 
with the feed water. Tliis claas of heater requires an efficient 
<h1 aeparatof oD the exhaust line to prevent cylinder oil being 
earned into the boiler. 

The Hcpai&tor is now being supplied with and made a 
part of the modem open type heater. The shell of the heater 
is either oonetructed of cast-iron ribbed platee or boiler plate 
and ia made either square, rectangular or round. In the latest 
type at open heater design a cutout valve is provided aa 
shown by Fig. 2. This arrangement obviates the necessity ^*^" 
at any by-paas piping around the heaters which requiiea 
the USB c£ thiee valves in order to cut out the beater for 
repaifs or cleaning. 

The upper part of the shell contains a ntunber of re- IH 
movable traye over which the incoming feed water trickles °^ 
and mingles with the exhaust steam, bdng heated to a tem- 
poatute of apiHuximately 200° to 210° when supplied with 
sufficient exhaust steam. Such scale forming matter as car- 
bonates of lime and magnesia, which will precipitate below 
this temperature, is depouted tm the trays. In the base 
of the heater a filter bed of charcoal ot coke is provided 
thioui^ wluch the water must paas on its way to the txMlers, 
for the further removal of such precipitate and impurities 
that can be removed by filtration. This type of heater is 
oftai rrfmed to as ttfeed water heater and purifier. A feed 
water metering device as a part of the heater may now be ^'"- ^■ 
obtained, the meter being of the V notch weir type with 
automatic reccffding device. Fig. 3. 

The adminion of feed water is ctmtrolled by a valve operated by a float located within the 
heater, an overflow pipe being provided to prevent flooding in case the float should fail to work. 

The feed pump in connection with the open heater must be located between the heater 
and boiler and should be placed a sufflcient depth below the heater to always be primed as the 
pomp must handle hot water. 

Unless the heater is supplied from a pressure main, an addilitmal pump will be required to 
draw water from the sump, hot well or other source of supply and deUver it to the heater. 

EaMtomy of Heating Feed Water. Economy due to feed wato' heating ia an important item 
ia power plant operation and for this reason few plants are constructed without heaters. Roughly 
^Making, f<v every 11 decrees the feed water is heated by exhaust steam othnwise wasted pro- 
duces a saving of 1% in the fuel required. This, however, does not take into account the initial 
east of heater, interest, depreciation, attendance and repairs that must necessarily follow any 
installation of apparatus to cause a saving. 

t) ■> n + 9i ~ heat content in one pound of dry saturated steam above 32° F. 
(i = temp, of the cold water. 
fi » temp, of water leaving heater. 
9i — heat of the liquid at (i° F. 
{t — beat of the liquid at fi° F. 



210 



POWER PLANTS AND REFRIGERATION 



The per cent saving in fuel is: 



5 = 100 X ^^' . = 100 X r (approximately). See Table I. 



(qt + rt) - ^1 



ts - ^l 



Example. Find the per cent saving due to a rise in temperature of the feed water from 60^ F. to 
210^ F., boiler pressure 160 lb. per sq. in. 

qt = 178 B.t.u. 
9i = 28 B.t.u. 

The heat content (i) in steam at 160 lb. per sq. in. pressure gage or 175 lb. absolute » 1195.9 B.t.u. 
Then 

178-28 ,^ ,„^^ 

^ = im:^^^ >< ^«» = ^^-^^^ 

or 1% for each 11.6** the feed water is raised in temperature. 

Final Temperature of Feed Water. The final temperature to which feed water may be 



Exhaust 
Inlet 




iler Feed Pump 



Pig. 3. DiAGEAM SHOWiNa Pbinciplb of thb Cochbane Oombinbd Open Peed Water Heater akd 

Meter. 



raised depends upon the amount of exhaust steam available for this purpose, the initial tern- 
perature of the feed water and the temperature of the exhaust steam. 

The temperature of the exhaust will dep)end upon the back pressure carried. With the 
open type of heater, the pressure in the heater is ordinarily atmospheric, corresponding to a 
temperature of 212^. Except in the case of a non-condensing plant in which the excess exhaust 
is used in a heating system, in which event it rarely ever exceeds 5 pounds, corresponding to a 



FEED WATER HEATERS AND FEED WATER PURIFICATION 



211 



temperature of &pproxiniatel7 227* F. In the cloeed type of heater, the final temperature that 
may be obtwned is lees for the aame back preoeure owing to the fact that the heat tranamiflMon 
of the tubee becomes impaired by scale and frequently an insufficient amount of heating surface 
■a supplied. The final temperature of the feed water with cloeed heaters ia rarely ever more 
than 200° with atmospheric exhaust and 210° with open beaten. 

Vbui Temperature of Feed Water hi a CoudensinK Plant Open HeaUr Supplied by Exhavat 
fnm lilt AustQiariea. An approximation for the quantity of steam used by the aunliaries can 
be taken as 10% of the total steam used by the main engines in the plant. The radiation loss 
and other kmea can be assumed as 10%. The final temperature of the feed water for any given 
ctnditkuta may be estimated as foUowa; 

Let (q + zr) » beat content above 32" in one pound ot the exhaust steam corresponding to 
the pressure maintained in the heater. 

gi — heat in the water entering the heater, per lb. 
qt " heat in the water leaving the heater, per lb. 
ID » wt. of exhaust (H' auxiliary steam available per hour, Va, 
X — quality of steam entering the heater. 
W — wt. of feed water used by main units- 
Heat entering heater above 32* - [Wq, +w(q + xr)] 0.9. 



Heat leaving heater above 3 



Approximately vt — (ft — 
Then 



' - (IT + w)5, 

{q+xr)v> + Wqi 



Hq +xr) iv +Wq,]X 0.9 



For preliniinary calculations the value of x may be taken as 1 for exhaust steam and w — 
aiOIF. 

TABLE' 1 

PERCBNTAGB OF 8AVIMO FOR E&CRJIBGRBE OF INCR£A3£ IN TBUPERATDBE OF FEED WATER 



p Btb4m is Boiua, Ls. r; 



POWER PLANTS AND REFRIGERATION. 



o-w*it. arrAff oart^r 



TYPE loo--^ooHP 



TrrE soo -730 HP. 



D WATKB HBATIBa. 



TABLE 2 

OPEN-TVPB FEED WATER HEATERS. <8mF1c4.) 

Dumnn of Pira CoNNKmoia, Etc 



poirer. 


Eihamt 
IniM 


Xihaurt 
Outlet 




Llva 


InlMfrom 




Ovwflow 


Blow-all 
udDr.ii> 


WsSf 




N 


o 


L 


Y 


X 


P 


H 


B 


s 


100 

400 
SOO 
TM 


s;; 

8" 

V- 


6" 

«" 

s: 

10" 


1" 

ZM" 


)' 


1 


2" 
6" 


K 

f 


if: 





DmBMStOHS OF HBATS 





A 


C 


D 


E 


F C 


H 


I 


J 


w 




1 


\ 


ill 

*' 9" 




1' 0" 

' 1" 


a" 

8" 


1 


IF 




III 





FEED WATER HEATERS AND FEED WATER PURIFICATION 



FiQ. 6. Opin Feed Watbb HMiraiia. 



TABLE 3 

OPEtt-TTFB FEED WATER BEATERS (ShPIc-G.) 
DuKBTSK or PiPB CotnncmiKi, Eia 



Bin*- 


"Sir 


OuUrt 


,»^ 




IS 


fe 


Overflow 


DniD 


ouXt 




« 





L 


Y 


X 


P 


M 


B 


3 




12" 








4" 




*" 




„ 












































































•»' 






•000 


S" 


la" 


*" 


... 


5" 


8" 


B" 


" 











DnmoiOHB 


or Hut 


ns 












A 


" 


D 


D, 


= 


F 


o 


H 


I 


J 


W 


™ 


w 


a'l" 


S'l" 


B- ■" 


B" 


fi" 






f!*- 


8' 7" 


7' 10" 


















;;j>" 


























%V^' 
























































0" 


" 




B'BM" 


..«" 





POWER PLANTS AND REFRIGERATION 

- ratio of the weight ot exhaust Bteam from auxiliaries per hour to the total wei^ 
ot feed water or steam eenerated by b<»len per hour 



(, - 0.9 [q + xt)R + (1 - R) 5,1 + 32. 
If 1 = 1, then I, - 0.9 [iR + (1 - fl) q,] + 82. 

Bxample, AiimmiDg ao initial tarn pen tun of feed water DO* P., (91 ~ 58); wn^t of ateam iwed 
hy auxiliariefl, w " 1,000 lb, per hour.. Wd^t of Bteam used by main unit, W — 10,000 lb. pa 

Then fi — =-; — = 0.091. 

W + « 

Aamning z - 1, then 1, - 0.9 [llSl.7 X 0.091 + (1 — 0.001} 66] + 32 - 171' F. 

Hie maximum Umpenivre to which the feed 
wat«r may be heated in an open heater ia 
apprcndmately 210° F. with atmospherie BxhausL 
With an initial temperature of feed water of 
60°, (gt - 2S) and I, - 210°, we find that B - 
0.15. That is, only 15% of the total exhaust ateam 
in a non-condenaing plant is necessary to heat the 
feed water to the nmTininm temperature with 60 de- 
grees feed water. Any excess exhaust above this 
amount is available for heating or prooeos work. 

BxhAust Stecm Available for Heating in Hon- 
condensing PUnts. In non-condensing pUnts whae 
the exhaust Eteam is used for low-prenure heating 
the back prewure carried on the heater should not 
OTdinarily exceed 5 lb. gage and with a properly 
designed vacuum system of beating the bade pre»- 
sure should not exceed 2 lb. gage. 

If the condensation from the radiatim is re- 
turned to the heato' a temperature of approxi- 
tr mately 150 degrees may be assumed for the initial 
temperature (1 of tiie feed water. This will <»di- 
narily provide tor the lowering of the temperature 
r of the condensate by the introduction of the oold 
makeup watw to ofTset the loss by leakage in tbe 
heating system. 

The weight of exhaust steam condensed in the 
feed water heater and the weight available for 
heating and process work may be estimated by 
the following formula: 
Let P •= weight of exhaust steam condensed in the feed water hoAxr pa hour. 
W + ID — P •= weight of exhaust steam available for heating per hour. 

y."^+ -""-'■' lb. 
0.9 (IT + s - ») 

^„._, „ (IT + »)«.-■,) , 1^^ 



FEED WATER HEATERS AND FEED WATER PURIFICATION 



215 



The percentage of the total weight of exhaust steam coxKlensed in the feed water heater per 
B>. of exhaust is 100/ per cent, and the percentage available for heating is 100 (1 — /). 

Szample. Afisume that a back pressure of 2 lb. gage ia carried on the heater and that the initial 
temperature of the feed water U » 150. For 2 lb. gage (16.7 lb. abs.) U = 219*»F., % « 1164. 

100 (219 - 150) ^ ^ _, 

/ ■■ na/^tiTA — t gn -L Qox ■ ''°/o* The percentage of the total ezhauBt available for heating, by 

weifl^t, is 100 - 7.8 » 92.2%. Practically a deduction of 5 to 10% should be made from the above 
figure to allow for the condensation in the steam mains to obtain the net weight available at the radiators. 

Specifications for Open Type Heater and Receiver with Provision for Purifying the Surplus 
Racluuist Steam Passing to the Heating System. The heater is to have ample capacity for heat- 
ing the water required for hp. of boilers, including such overload as may be carried on 

the Ixnlers, taking the initial supply at 50° F. and delivering it at a* temperatm^ within from 
2 to 5** ol the temperature of the steam entering the heater, when the heater is kept filled with 
steam. 

TABLE 4 

DIMENSIONS OF THE VERTICAL OTIS CLOSED TYPE HEATER 



Number of Heater 
(SiM of Exhaust) 



4 

A4 

B4 

C4 

6 

A6 

B« 

7 

A7 

8 

AS 

9 

A9 

10 ^ 

AlO 

BIO 

12 

A12 

B12 

16 

A16 

B16 

18 

A18 

B18 

C18 

D18 



Horse- 
power 



80 

40 

60 

60 

100 

125 

150 

160 

200 

250 

800 

850 

400 

450 

500 

550 

600 

700 

800 

900 

1,000 

1400 

1450 

1^50 

1.400 

1.500 

1,700 



Siaein Inches 



Number 

of 
Tubes 



15 
15 
15 
15 
20 
20 
20 
25 
25 
25 
25 
80 
30 
85 
85 
85 
40 
40 
40 
45 
45 
45 
54 
54 
54 
54 
54 



48 
60 
72 
84 
72 
84 
96 
72 
84 
96 
zl08 
zl08 
zl20 
zl20 
zl82 
zl44 
zl82 
zl44 
zl56 
zl44 
zl56 
zl68 
zl44 
zl56 
zl68 
zl80 
zl92 



14 
14 
14 
14 
24 
24 
24 
48 
48 
48 
48 
52 
52 
52 
52 
52 
60 
60 
60 
56 
56 
56 
86 
86 
86 
86 
86 



Sq. Ft. 
Heating 
Surface 



16 

28 

80 

86 

68 

66 

80 

88 

105 

160 

170 

176 

205 

225 

257 

290 

819 

861 

408 



494 
551 
568 
643 
720 
796 
872 



Dia.Fted 

Pipe, 

Inches 



1 
1 
1 
1 

1^ 
1 

i; 

2 
2 
2 
2 

2H 

2H 

8 

8 

8 

8] 

8! 

4 
4 
4 
5 
5 
5 
5 
5 



Weight 



660 

620 

680 

740 

1,090 

1,170 

1,260 

1,540 

1,670 

1.790 

1,920 

2,480 

2,680 

3,400 

3,600 

8,800 

6.500 

6.750 

7.000 

9.400 

9.800 

9,900 

14,400 

14,900 

15,400 

15,900 

16,400 



Norm. — ^The horizontal types have same diameter as the votical, but are a few inches shorter. The number given 
to the heater & the table is the largest diameter of exhaust pfpe the heater is adapted for. The heating surface given 
ia the actual heating surface ot the tubes and water separator. 



Tlie heater is to have a water storage capacity below overflow level of not less than 

cu. ft. With the heater is to be furnished an oil separator of approved design (self-cleaning 
type) and of ample capacity for purifying exhaust steam to an amoimt equivalent to the full 

rated capacity of the heater, namely boiler hp. Also, such trap or traps as may 

be necessary for draining the oil separator and taking care of the overflow from the heater, the 
valve area of the trap to be not less than the full area of drip pipe from the separator, that is, 
the area of the valve in the steam trap is to be not less than the area of a inch pipe. 

The heater is to be a unitary structure comprising a heater and separator, and means for 
controlling the passage of steam between the separator and the heater, all so arranged that the 
heater can be isolated or cut off, for examination or cleaning, from the path of steam to the 



216 



POWER PLANTS AND REFRIGERATION 



heating system or to atmosphere. The separator is to continue in operaticm when the 
heater is cut out, at which times the drainage of the separator is to continue independently of the 
overflow drainage. 

The heater is likewise to be provided with readily removable cast-inm trays, oold-water 
regulating valve and float for controlling the admission of the cold water supply under a pres- 
sure on the cold water supply line of from 10 to 30 lb. Suitable provision is also to be made 
so that filtering or depositing material may be carried within the heater under downward fihn- 
tion. Pump supply is to be hooded and vented to steam space. 

Closed Heaters. A closed heater consists of a circular shell in vihlch are placed a number 
of straight or curved tubes, usually seamless brass. If the exhaust steam surrounds the tubes, 
the feed water passing through the tubes, it is known as a woUer-tube type of closed heater. If 
the reverse is true, it is known as a sUamAube heater. 

The closed type of feed water heater is sometimes employed in a condensing plant by plac- 
ing it in the exhaust line between the engine or turbine and the condenser. When used in this 
connection it is frequently termed a primary healer or vaeuttm heater. The feed water after 
having passed through the primary heater is delivered to either an open or closed type of 
heater, to ^diich the exhaust from the auxiliaries is delivered. The temperature to which 
the water may be raised in the primary heater will be approximately 10** lower than the tem- 
perature of the exhaust steam, which for a 26" vacuum is 116** F. The final temperature of the 
water leaving the primary heater will probably not exceed 105^ for this degree of vacuum. The 
closed heater is also used for heating purposes in connection with forced hot-water circulating 
systems, as described in the Chapter on ''District Heating,'' Volume I. 

As the steam and water are never in direct contact with one another, the eflicienoy of this 
type depends upon the amount of heating surface and its conductivity. 

Closed heaters are often spoken of as having a rated horsepower. This is a conmiercial rating 
and is usually based on ^ sq. ft. of heating surface per boiler horsepower. Heaters should not 
be purchased on this rating, but upon the sq. ft. of heating surface required to transmit the 
necessary amount of heat to raise the temperature of the feed water a given amount. With 
the closed type of heater one feed pump only is required. 

The size of a closed heater vri]! depend upon the rate of conductivity of the metal used in 





Fig. 7. DncKNBiONs of Olobkd Hkatbbs. (See Table 6.) 



FEED WATER HEATEBS AND FEED WATER PURIFICATION 



217 



t 

I 



o 
n 



«< 



»s 



it 



o 

00 

Z 
O 

g 



Bo-iftO|S 








H 


1 






























M 










» 


«j«a 


:s! 


•H 


f4 


a; 


M 


C4 


CO 




^ 


:st 


f4 






1 


nrm 


» 


55 

ot, 


ot. 


o!. 


noa 






0» 


o 




poaaog 


o 






WIS 


v4 


3; 


N 




00 


^ 


le 


<0 


1 
1 


■Hoa 




X 

^ 


4. 


ot, 


oi, 


oi, 


•1M!0 

»ioa 




t- 




00 




^4 


S5S 






PM^ 


• 


3; 

00 


a* 


o 


f4 


09 

^4 


«|S 




M 


C4 


09 


09 


•^ 


lO 


<0 


00 


5 

B 

H 


■pc^S 






3^ 

ot, 


ol, 


ob 


e!. 


e!. 


^4 


^4 




f4 
C4 


9p^o 






3; 

00 










00 




:s? 


^ 


aBaiu 




0» 


o 




09 




0» 






to 

C4 




e«!S 


09 


^ 


le 


<0 


00 


o 




»H 




00 


S 


A 


<D 


t- 


00 

• 




e 

^4 


^4 






:« 
SS 






Si 




00 


t- 


00 




e 

^4 


:5! 

^4 


CO 

^4 


^4 


09 






A 






^ 




<0 


3; 

00 




^4 








lO 


^4 


A 




;:j^ 
n 


«D 




a* 




^4 


^4 




to 


35 

e 




kO 


;s? 






lO 


Ok 






^4 


CO 

^4 




lO 




S 




:« 
S 


^ 


s 


' 






O 




3! 




O 
CO 


S 


:« 
5 


s 


lO 


09 




s 






00 


s 


S 




o 




:« 

S 


:« 
^ 


<0 


M 






9 


9 


S 3 


;« 
? 


:« 
? 




3 


g 


s 


e 


? 


s 


•ON 






;5 


is? 










^ 


0) 




9 


8 






s 


S 


;« 

sg 




U 


3 


3! 

s 




;« 
S 


« 

s 


00 


So 




9-t 
O 


^ 

s 


kO 
0» 


;s5 

e 

^4 


kO 


:s? 

00 


Rating 

in 
B.Hp. 


s 




s 


§ 


s 


i 


§ 


i 


3 

CO 


i 


i 


i 


i 


1 


i 






»H 


wt 


1 




ea 


S 

3 


* 

CO 




^ 


<0 




1 


A 


s 


4 


.8 


d 


.S 


<S 


s 


f4 




S 


S 


3 








f4 




1 




a 


S 


1 




i 


g 


a 



POWER PLANTS AND REFRIGERATION 



jl l\ 

a iiJ tad' 

ill lii 



FEED WATER HEATERS AND FEED WATER PURIFICATION 



219 



the tubes, vdiich in turn is dependent upon the rate of flow of the water and the number of passes 
the wat^ makes throu^ the heater. 

Tlie conductivity taken from experiments by various authorities has been found to be ap- 
proKiniately as folkyws: 

Let U » B.t.u. transmitted to the feed water per sq. ft. of surface per hour per degree 
. difference in the average temperature of the steam and feed water. 
Average vahies iox U: 

MtdHpU-'flaw heaters, vek>oity of water 50 ft. per minute. 

Plain copper tubes 250 B.t.u. 

Conugated copper tubes 300 B.t.u. 

Single^flow heaters, velocity of water 12.5 ft. per minute. 

Plain brass tubes 175 B.t.u. 

Coil^pe heaters, velocity of water 150 ft. per min. 

Plain copper tubes 300 B.t.u. 

For steam-heated tubes 

Plain iron tubes 120 B.t.u. 



TABLE 8 

GENERAL DATA AND APPROXIMATB NET SELLING PRICES FOR FEED WATER HEATERS 

Open Type — ^For power planti operating steftm-heAtfais vyeteme 



HiWB^pQ'WP ISiUUf • •• •■••••■••••■••■•• 


60 

1600 

1660 

1186 

84 

26 

67 

i^ 

IV- 

17 
11 


76 

2260 

1760 

166 

86 

27 

71 

6 

1 

2 

^1 

4 

19 
18 H 


100 

8000 

1860 

188 

88 

29 

76 

6 

1 

2 

2 

2 

4 

21 

16 


160 

4600 

2700 

289 

41 

82 

66 

7 

2H 

2H 

6 

22 

16 


200 

6000 

8000 

290 

48 

84 

76 

8 

^1 

2H 

8 

6 

24 

16 H 


800 

9000 

4800 

848 

48 

88 

81 

9 

2H 

4 

6 

28 

21 


460 

18500 

6860 

440 

62 

48 

87 

10 

2 

4 

8 

4 

10 

82 

10 H 


600 

18000 

6760 

646 

66 

47 

98 

12 

2 

10 
86 
12 


900 

27000 

8160 

696 

86 

48 

87 

14 

^1 

4 

6 

20 

82 

10 H 


1200 

86000 

11000 

866 

94 

47 

98 

16 

8 

6 

*n 

20 
86 
12 


1600 

48000 

12000 

1060 

96 

68 

98 

18 

8 

6 

4H 

7 

40 

21 

12 


2000 
60000 
18000 

1190 
98 

6 

J 

24 

12 


2600 




^6000 
14600 


roonde of leea wucr neeiea per noinr. . . . 
Wiifht fai poinide 


Net'briee fl Oi. b. fMtonr 


1880 


Wm5», inchw 


100 


I>Hth,toelHe 


67 


H«(^t» fawhw 


93 


I>fauiBe.esh.iBllBt end oatlat^^eny itwupto 
Die. ine. eold water mxpfiy 


20 

"1 




Dia. ine. wnato asd overflow 


6 




8 


rf wnhflT Qi tmys . .•*.......«••..•.•.•• 


40 


lenctii Der txiTy. Inchoi 


28 


WioUi Der tnur. indiee 


12 







Theee beatere, wUle performing all the fnnetione of an opm lieater for the power plant, are alao derigned to 
9df and to heat the eondeiMation returned from a iteam-heating ayetem. It will be noted that this double eervice 
reipiiraa a aooiewhat Ivfer beater than that reqidred for the aenriee of the following table. 

Priew ai« net f. o. h. New York <^ty and may be aafely need for eetimating and yaluation p ur poee e . To cover 
fnight* add 76 eenta per hnndredweii^t for every 1,000 milee from New York. 

To edeet the proper heater tnm thia table, utiliae the a pp roximation method outlined below following taUe, but 
add to the amount determined the water reouired for the iteam-heating ajretem. Consider that the aiae of the heater 
b gowned by the total quantity of water wUdi must be peaeed through the heater in a given time. Consider, further, 
ttat wUle the water in ue eteam^heating syatem is praeoeallya constant quanti^in continuous dreulation,it ii 



to 



bom the 



of leakage^ evaporadbn, drain-off, etc Theee I 



is subject 
must be made up by the supfdy oi additional 



In dosed heaters the temperature of the water and the temperature of the steam can never 
be equal, and for practical purposes may be taken as ^s » f« — 10, where 1$ » temperature of 
steam and it « temperature of water leaving the heater. 

Let ti » temperature of water entering the hea^ter. 
(a s temperature of steam entering the heater. 
it -> temperature of water leaving the heater. 
A B sq. ft. oi transmitting surface. 

d -> mean difference in temperature between steam and feed water. 
TT -> wt. of feed water heated per hour in lb. 



220 POWER PLANTS AND REFRIGERATION 

Then AUd^ TT («i - « 

W(tt - k) 



A = 



Ud 



d = is — approximate, but near enough for practical purposes. 



Then A = 



W{tt - ti) 



" ('• - H^i 



FEED WATER PURIFICATION 

Natural waters all contain some impurities, which are either soluble, insoluble or both« 
The impurities are divided into two general classes, incrusting and non-incrusting. The former 
is composed principally of the lime and magnesia salts and aU suspended matter; the latter 
includes only the sodium salts. When the water is evaporated into steam aU of the impuri- 
ties, including the suspended matter, is left in the boiler. After a period the concentration 
becomes so great that the scale-forming impurities crystallize and are deposited in the boiler 
along with the suspended matter in the form of sludge or scale. 

The effect produced on the boiler, if the impurities are not removed, is: (1) a reduction 
in the heat transmission of the boiler heating surface and, therefore, a reduced steaming capacity 
and fuel waste; (2) the liability of overheating the tubes and plates, thus producing a dangerous 
condition of operation. The salts usually responsible for incrustation are the carbonates 
and sulphates of lime and magnesia, and boiler feed treatment in general deals with tiie 
getting rid of these salts more or less completely. The table on page 221, by W, W, Christie^ 
gives an approximate classification of impurities found in feed waters, their effect and the rem- 
edy or means for overcoming the effect produced. 

Treatment of Feed Water. The following matter has been taken, in part, from " Steam " 
(Babcock and Wilcox Co,), 

Scale Formalion, The treatment of feed water carrying scale-forming ingredients is along 
two main lines: 1st, by chemical means by which such impurities as are carried by the water 
are caused to precipitate; and, 2nd, by the means of heat, which results in the reduction 
of the power of water to hold certain salts in solution. The latter method alone is suffi- 
cient in the case of certain temporarily hard waters, but the heat treatment, in general, is 
used in connection with a chemical treatment to assist the latter. 

Before going further into detail as to the treatment of water, it may be well to define certain 
terms used. 

Hardness, which is the most widely known evidence of the presence in water of scale-forming 
matter, is that quality the variation of which makes it more difficult to obtain a lather or suds 
from soap in one water than in another. This action is made use of in the soap test for hard- 
ness described later. Hardness is ordinarily classed as either temporary or permanent. Tem- 
porarily hard waters are those containing carbonates of lime and magnesium, which may be 
precipitated by boiling at 212^ and which, if they contain no other scale-forming ingredients, 
become " soft " imder such treatment. Permanently hard waters are those containing mainly 
calcium sulphate, which is only precipitated at the high temperatures found in the boiler itself, 
300° F. or more. The scale of hardness is an arbitrary one, based on the number of grains of 
solids per gallon, and waters may be classed on such a basis as follows: 1-10 grains per gallon, soft 
water; 10-20 grains per gallon, moderately hard water; above 25 grains per gallon, very hard water. 

AlkaUnity is a general term used for waters containing compounds with the power of neu- 
tralizing acids. 

Causticity f as used in water treatment, is a term coined by A, McGiUy indicating the prcs- 



FEED WATER HEATERS AND FEED WATER PURIFICATION 



221 



eooe ol an excess of lime added during treatment. Thoi^gh such presence would also indicate 
alkalinity, the term is arbitrarily used to apply to those hydrates whose presence is indicated 
by pheno^hthalein . 

Chemical Treatment Of the chemical methods of water treatment, there are three general 



1st. lAme Process, The lime process is used for waters containing bicarbonates of lime 
and magnesia. Slaked lime in solution, as lime water, is the reagent used. Thb combines 
with the carixmic add which is present, either free or as carbonates, to form an insoluble mono- 
carb(mate of lime. The soluble bicarbonates of lime and magnesia, losing their carbonic acid, 
thereby become insoluble and precipitate. 

2iid. Soda Process, The soda process is used for waters containing sulphates of lime and 
magnfflia. Carbonate of soda and hydrate of soda (caustic soda) are used either alone or to- 
gether as the reagents. Carbonate of soda, added to water containing little or no carbonic acid 
or bicarbonates, decomposes the sulphates to form insoluble carbonate of lime or magnesia which 
precipitate, the neutral soda remaining in solution. If free carbonic acid or bicarbonates are 
present, bicarbonate of lime is formed and remains in solution, though under the action of heat 
the carbon dioxide will be driven off and insoluble monocarbonates will be formed. Caustic 
soda used in this process causes a more energetic action, it being presumed that the caustic soda 
absc^foe the carbonic add, becomes carbonate of soda and acts as above. 



Trouble 



Inonistatioo. 



Priming, 



Cauae 



Sediment, mud, day* etc, 
Readily aoluble aalta 



Bicarbonate of magneala, 
lime, iron 



Organic matter 

Sulphate oi lime 

Organic matter 

Greaae 

Chloride or sulphate of 

magnedum 

Sugar 

Add 

DiMolTed carbonic add and 
oxygen 

Electrolyt i c action 

Sewage 

Alkdiea 

Carbonate of aoda in large 
quantitiea 



Remedy or Palliation 



Ffltration. 

Blowing off. 

Blowing off. 

Heating feed and predpitate. 

Caustic soda. 

lime 

Magnesia. 

See organic matter under corrosion. 

Sodium carbonate. 

Barium diloride. 

Predpitate with alum , ..^ m^mm 

Predpitate with ferric chloride / "** "*•• 

Slaked lime \ ..^ mt^ 

Carbonate of aoda / ««»fflter. 

Carbonate of soda. 
AlkalL 

Slaked lime. 

Caustic soda. 

Heating. 

Zinc plates. 

Predpitation with alum or ferric diloride and filtv. 

HeaUng feed and predpitate. 

Barium diloride. 



3rd. Lime and Soda Process, This process, which is the combination of the first two, 
is by far the most generally used in water purification. Such a method is used where sulphates 
of lime and magnesia are contained in the water, together with such quantity of carbonic acid 
or bicarbcHiates as to impair the action of the soda. Sufficient soda is used to break down the 
sulphates of lime and magnesia and as much lime added as is required to absorb the carbonic 
add not taken up in the soda reaction. 

All of ihe apparatus for effecting such treatment of feed waters is approximately the same 
in its chemical action, the numerous systems differing in the methods of introduction and handling 
of the reagents. 

Heat Treatment. Sediment, mud, day and all suspended matter may be removed from 
feed water by filtration. The materials ordinarily used for the filter are ook6 and excelsior. 



222 POWER PLANTS AND REFRIGERATION 

Some of the seale-fomimg matter held m solution, such as bicarbonate of magnflnia and lime, 
may be removed by precipitation by frst heating the feed water. 

The modem open type heater will heat the feed water to i4)proximately 210^ F. and such 
scale-forming substances as are precipitated below this temperature are deposited on the trays 
and in the settling chamber. Tlie sulphates of lime and magnesia require a temperature from 
290** to 300** F. for complete precipitation and therefore will not be completely removed by an 
open heater. 

live steam heaters are used for purifying feed water containing the sulphates of fime and 
mftgnflwift. alone or in connection with the bicarbonates. The usual type is fitted with remov- 
able trays. The water to be purified discharges into tne upper pans and overflows into the 
lower pans and to the lower part of th^ heater from which the feed water is drawn. The purifier 
should be located about two feet above the level of the boOer water line so that the feed water 
will flow by gravity into the boilers. live steam purifiers when used are ordinarily opaaAed in 
conjunction with exhaust steam heaters. 

An economizer will also precipitate the sulphates of lime and magnesia when the maintained 
temperature is 290** F. and above. 

Combined Heat and Chemical Treatment. Heat is used in many systems of feed treatment 
a{^>aratus as an adjunct to the chemical process. Heat alone will remove tempcHary hardness 
by the precipitation of carbonates of lime and magnesia and, when used in connection with the 
chemical process, leaves only the permanent hardness or the sulphates of lime to be taken care 
of by chemical treatment. 

The chemicals used in the ordinary lime and soda process of feed water treatment are com- 
mon lime and soda. The efficiency of such apparatus will depend wholly upon the amount and 
character of the impurities in the water to be treated. Table 9 gives the amount 6i lime and 
soda required per 1000 gallons for each grain per gallon of the various impurities found in ibe 
water. This table is based on lime containing 90 per cent calcium oxide and soda containing 
58 per cent sodium oxide, which correspond to the conmiercial quality ordinarily purchasable. 
From this table and the cost of the lime and soda, the cost al treating any water per 1000 gaUona 
may be readily computed. 

Less Usual Reagents. Barium hydrate is sometimes used to reduce permanent hardness 
or the calcium sulphate component. Until recently, the high cost of barium hydrate has ren- 
dered its use prohibitive, but at the present it is obtained as a by-product in cement manufacture 
and it may be purchased at a more reasonable figure than heretctfore. It acts directly on the 
soluble sulphates to form barium sulphate, which is insoluble and may be precipitated. Where 
this reagent is used, it is desirable that the reacti<Hi be allowed to take place outside of the boiler, 
though there are certain cases where its external use is permissible. 

Barium carbonate is sometimes used in removing calcium sulphate, the products of the 
reaction being barium sulphate and calcium carbonate, both of vihlch are insoluble and may be 
precipitated. As barium carbonate in itself is insoluble, it cannot be added to water as a solu- 
tion, and its use should, th^^ore, be confined to treatment outside of the boiler. 

Silicate of soda will precipitate calcium carbonate with the formation of a gelatinous silicate 
of lime and carbonate of soda. If calcium sulphate is also present, carbonate of soda is formed 
in the above reaction, which in turn will break down the sulphate. 

Oxalate of soda is an expensive but efficient reagent which forms a precipitate of cakium 
oxalate of a particularly insoluble nature. 

Alum and iron alum will act as efficient ooagulents where organic matter is present in the 
water. Iron alum has not only this property but also that of reducing oil discharged from surface 
condensers to a condition in which it may be readily removed by filtration. 

Corrosion, Where there is a corrosive action because of the presence of add in the water 
or of oil containing fatty acids whicfe will decompose and cause pitting wherever the sludge can 
find a resting place, it may be overcome by the neutralization of the water by carbonate of soda. 
Such neutralization should be carried to the point where the water will just turn red litmus paper 



FEED WATER HEATERS AND FEED WATER PURIFICATION 



223 



Uue. Ab a jn^ventive of such action arising from the presence of the oil, only the highest 
grades of hydrocarbon oils should be used. 

TABLE 9 

REAGENTS REQUIRED IN UME AND SODA PROCESS FOR TREATING 1000 U. S. GALLONS OF WATER 

PER GRAIN PER GALLON OF CONTAINED IMPURITIES* 





liBMt 

Pounds 


Sodai 
Pounds 




Timet 


Sodai 
Pounds 


Oridnm Cwbonato 

Odctam 8gli*****?B 

Ctfcfam Chkrids 


0.098 

• • • • . 

• • . • . 
.284 
.079 
.108 
.007 


• • • • • 

0.124 
.161 
.104 

• • • • • 
.141 
.177 
.115 


Fvroos Carboiiate 

FsrrouB Sulphate 

Ferrie Sulphate 

Aluminum Sulpliate 

Ftm Sulphurie Add 

Sodium Carbonate. 

Ftm Carlnm Dioxide 

Hydrofen Sulphite 


0.169 
.070 
.074 
.087 
.100 
.098 
.228 
.288 


• • • • • 

0.110 
.126 


Cydnm Nitrate 

Mifimdnin Chloride 

MagMiiam NitnUa 


.147 
.171 



^L,M. BoaA Cosiponf . 

' Based on lime rontelning 90 per eent caldum oxide. 
Di— d on aoda containing 68 psc mat sodium oiide. 



\ 



Addity will occur where sea water is present in a boiler. There is the possibility of such 
an occurrence in marine practice and in stationary plants using sea water for condensing, due 
to leaky condenser tubes, priming in the evaporator^ etc. Sudh acidity is caused through the 
dissociation of magnesium chk^de into hydrochloride add and magnesia under high tempera- 
tores. Tlie add in contact with the metal forms an iron salt whidi inunediately upon its for- 
mation is neutralised by the free magnesia in the water, thereby predpitating iron oxide and 
reforming Trm gnflwiiim chlcmde. The preventive for corrodon arising from such addity is the 
kee|»ng tig^t of the condenser. Where it is unavoidable that some sea water should find its 
way into a^bcnler, the addity resulting should be neutralized by soda ash. This will convert 
the magnesium chloride into magnedum carbonate and sodium chloride, ndther 6i which is 
COfio d ve but both of which are scale-forming. 

The presence c^ air in the feed water which is sucked in by the feed piunp is a weQ-recognised 
oanse of oorrosicm. Air bubbles (orm bebw the water line and attack the metal of the boiler, 
the oxygen of the air causing oxidisation of the boiler metal and the formation of rust. The 
partide of rust thus f(»ined is swept away by the circulation or is dislodged by expandon and 
the minute pit thus left f<»ins an ideal resting place for other air bubbles and the continuation 
of the oxidisation process. The prevention is, of course, the removing of the air from the feed 
water. In marine practice, where there has been experienced the most difficulty from this source, 
it has been found to be advantageous to pump the water from the hot well to a filter tank placed 
above the feed pump suction valves. In this way the air is liberated from the surface of the 
tank and a head is assured for the suction end of the pump. In this same class of wwk, the 
ooiTodve action of air is reduced by introducing the feed through a spray nozzle into the steam 
space above the water line. 

Galvanic action, resulting in the eating away of the boiler metal through dectrdysis, was 
formerly conddered practically the sole cause of corrodon. But little is known of such action 
aside from the fact tiiat it does take place in certain instances. The means adopted as a rem- 
edy is usually the installation of zinc plates within the boiler, which must have podtive metallic 
contact with the boiler metal. In this way, local electrolytic effects are overcome by a still 
greater electrolytic action at the expense of the more podtive zinc. The podtive contact nec- 
essary is difficult to mftint *"" and it is questionable just what efficacy such plates have except 
for a short period after thdr installation when the contact is known to be podtive. Adde from 
protection horn such dectroljrtic action, however, the zinc plates have a distinct use where there 
is the liability of air in the feed, as th^ offer a substance much more readily oxidized by such 
air than the metal of the boiler. 



224 



POWER PLANTS AND REFRIGERATION 



Foaming. Where foaming is cauaed by organic matter in auspeoaoo, it may be luKcIj 
overcome by filtration or by the use of a coagulant in connection with filtration, the latter oon>- 
bination having come recently into considerable favor. Alum, or potash alum, and iron slum, 
which in reality contains no alumina and should rather be called potassia-femc, are the ocwgu- 
lanta generally used in connection with Bltrstion. Such matter as is not removed by filtration 
may, under cert^n conditions, be handled by surface blowing. In some instances, settling tanka 
are used for the removal of matter in suspension, but where large quantities <rf water are reciuiivd 




filtratimt is ordinarily substituted on acco\mt of the time element and the large area a 
in settling tanks. 

Where foaming occurs as the result of overtreatment of the feed water, the obvious remedy 
is a change in such treatment. 

Priming. Where priming is caused by excessive concentration of salts witlun a boiler, 
it may be overcome largely by frequent blowing down. The d^ree of concentration allowable 
before priming will take plaix varies widely with conditions of operation and may be definite 
determined only by experience with each individual set of conditions. It is the preeence <A the 
salts that cause priming that may result in the absolute unfitness o! water for boiler feed pur- 
poses. Where these salts exist in such quantities that the amount of blowing down necessary 
to keep the d^ree of concentration below the priming point results in excessive losses, the only 
remedy is the securing of another supply of feed, and the results will warrant the changa almost 
regardless of the expense. In some few instances, the impurities may be taken care of by scHue 
method <}£ water treatment, but such water should be submitted to an authority on tbe subject 
before any treatment apparatus is installed. 

Boiier Compoundi. The method of treatment of feed water by far tlie most generally UBad 
is by the use of some of the so-called boiler compounds. There are many reUable concerns ha»> 
dling such c<napounds who unquestionably secure the promised results, but there is a great Ma- 



FEED WATER HEATERS AND PEED WATER PtlftlFlCATION 2^5 

(kncy towmrd looking on Uie owipound aaa." aure-eJl " for any mter difficulties, and care should 
be taken to deal only with reputable conoerna. 

The compomtion of these compounds is almost invariably baaed on goda with certain tannic 
Bubstaooes, and in some instances a gelatinous substance which is presumed to encircle scale par- 
ticks and prevent their adhering to the bc^ltx surfaces. Th» action of these compounds is or- 
dinarily to reduce the calcium sulphate in the watec by means^f carbonate of soda and to precip- 



Fio. 10. ABHAMmHCin' OP Watbb Softbnwo Pburr. 

itate it as a muddy form of calcium carbonate which may be blown off. The tannic compounds 
are used in connection with the soda with the idea of introducing organic matter into any scale 
already formed. When it has peiketrated to the boiler metal, decomposition of the scale sets in, 
cauong a disruptive effect which breaks the scale from the metal sometimes in large ekim. It is 
this effect ot bcHler compounds that is to be most carefully guarded against or inevitable trouble 
will result from the presence of loose scale with the consequent danger of tube losses through 
ming. 
When proper care is taken to suit the compound to the water in use, the results secured are 



1 



226 



POWER PLANTS AND REFRIGERATION 



fairly effective. In general, howeveri the use of compounds may only be recommended for 
the preventimi of scale rather than with the view to removing scale which has already formed, 
that isi the compounds should be introduced with the feed water cmly when the boiler has been 
thoroughly cleaned. 

Water Treating Apparfttus. Boiler compounds are ordinarily introduced into the feed water 
in the suction line to the feed puiftp as indicated by Fig. 8. A H-in, nipple is connected into 
the suction pipe of a pump, on the end of which is an angle valve canying a short vertical nipple, 
followed by a reducing coupling. A piece of larger pipe, 2 ft. or more in length, is screwed into 
this coupling and a cap completes the device and gives a finished appearance to the feeder. If 
the pump takes water from a well or a brook there is a partial vacuum in the suction pipe; there- 
fore, when the angle valve is opened the atmospheric pressure forces the compound into the 
suction pipe. Where the supply comes under pressure through the vertical pipe, it is necessary 
to locate a valve below the feeder, as shown, and this should be partially closed when the ccun- 
pound is to be drawn into the suction or supply pipe. 

Fig. 9 illustrates a more elabcntite device for the same purpose, to be used in connection 
with a vertical pipe. The body consists of a piece of pipe large enough to hold the required 
quantity of compoimd after it is dissolved or otherwise prepared for use. It is assumed that 
this is larger than the connecting pipes, which ought to be equal to the suction pipe, ther^ore 
a reducing coupling will be required for each end, followed by a nipple and a cross valve, with 




Pio. 11. Booth Watbb Softensb with Storage Spacb. Type " P-14." 



FEED WATER HEATERS AND FEED WATER PURIFICATION 



227 



TABLE 10 

TYPE "F-14'' BOOTH WATER SOFTENERS 



Capadtgr OalloaB par Hoar 



MM. 
6,000 

10,000. 

Sfoo. 

15,000. 



Ht. 



12'-0" 


26'-0" 


ly-*" 


2e'-9" 


14'-6" 


2e'-9" 


ly-o" 


2e'-9" 


16'-9" 


26'-9" 


ly-y' 


2r-s'' 


21'-0" 


27'-«'' 


28'-0" 


27'-«" 


2e'-6" 


27'-«" 



Sg. 



2,690 
8.090 

i;7oo 

4370 
4,940 
6,190 
7,770 
9,820 
12370 



Wf. 



76 
98 
110 
129 
146 
181 
228 
272 
869 



Price 
F. O. B. 
Factory 



$1,960.00 
2,200.00 
2,600.00 
2,800.00 
8,160.00 
8,660.00 
4,160.00 
4,600.00 
6,260.00 



Bdght ot main »*«v, 19^-6". 

Hd^ of outlet at bottom of soft water etorace apaee. 16'-6' 

Sf— Solt wmtar atorace capacity above outlet in galkma. 



Boiler h o rae p o w er « capacity -l- 4. 
Tanka ahipped Imocked down. 

Wf— Wdgbt fined witb water in tona. 



anoUier valve in the main suction pipe, as shown. This device is operated as follows: When it 
is to be filled dose valves 2 and 3 and open valve 4. Valve 5 should be opened to drain out 
any water that may remain in the body and then closed again. If the pump is running valve 
6 most be open. The device should be filled through the funnel 7, then valve 4 dosed, valves 2 
tad 3 opened, and valve 6 closed in the order mentioned. 

Water Treating PlanU. There are virtually two styles of water treating plants: the inter- 
mittent open and the ocmtinuous open. These systems are operated with either warm or cold 
water. As chemical reacticm will take place more rapidly with warm water, consequently this 
style of plant may be smaller than when cold feed water only is treated. Either style of plant, 
owing to its siie, is ordinarily located outside of the power plant. Fig. 10 shows, in outline, a 
typical continuous open type water softener designed to handle 8,000 gals, per hour. (Engineer^ 
wg New$, Aug., 1910.) 

The water softener consists of a steel settling tank 14|^ ft. in diameter and 30 ft. high, 
with an iron stairway encircling it from ground level to top. In the top of the settling tank is 
a small rectangular tank ocmtaining the automatic measuring and mixing mechanism for adding 
the softening solution (lime and soda ash) to the crude water. As a part of this apparatus there 
b on top of the 30^. tower a semicircular tank, alwa3r8 containing about 40 gals, of solution, 
in the bottom of which is a valve operated by a lever and cam system connected with a tilting 
backet. 

The tilting bucket is located just beneath the solution tank and has two separate compart- 
ments. The crude-water pipe has its discharge just above the tilting bucket so that the crude 
water flows into one compartment until reaching a certain level Then the bucket becomes 
^mhalanced, dumps and brings the other ccMnpartment imder the crude-water discharge. The 
tilting is again caused in the reverse direction. This rocking works the cam and lever system 
of the chemical-feed valve so that the requisite chemicals are automatically added. At the same 
time the contents of the chemical tank are agitated. 

The crude water, after receiving the proper amount of softening solution, is thoroughly 
stirred by a paddle attached to the tipping bucket. With a whirling motion, given to it to hasten 
the coagulation of predpitated impurities, the water passes through a downtake to the bottom 
of the settling tank, from which point it gradually rises to the top of the tank. Fen: an added 
precaution, it is passed through an excelsior or quartz filter from which it flows by gravity to 
the storage tank. 

At the base of the settling tank are located air-tight bins for a month's supply of chemicals, 
a chemical-mixing and lime-slaking tank, and a specially designed, positive type of water motor 
that utilises the crude water for furnishing all the power needed for mixing up the chemicals, 
for keeping the solution constantly stirred while the plant is in operation and running the pump 
that ddivefB the solution from a tank at ground levd to the semicircular one at the top o( the 



Nora. Tor tint daaa remlta the velocity of the water in the aetUing tank should not exceed (r m 4 to 6 ft. 
Mr boor. 



228 



POWER PLANTS AND REFRIGERATION 



softener. A double pipe line connects these two last-mentioned tanks so that the surplus solu- 
tion required in keeping a constant level in the upper tank may continuously overflow to the 
lower tank while the plant is operating. The motor requires no free discharge and is operated 
only by the crude water going to the top of the softener for treatment. The excess water pres- 
sure required to operate the motor imder full load is from 2}^ to 3 lb. — equal to pumping the 
water an additional height of 6 ft. 

The main solution tank on the ground level is designed to contain sufficient sc^tening solu- 
tion for 12 hours' continuous operation, making attendance oftener than once in that time un- 
necessary. Opening the water valve at the motor places the entire plant in inmiediate and 
automatic operation. 

This plant was given a thorough test both in efficacy and in economy of operation. It was 
found that repeated analyses showed a practically unchanging quality of softened water, and 
copies of t3rpical analyses given in Table 11 testify to the excellent degree to which the softening 
process is carried. From analysis No. 2 it is seen that the removal of scale-forming impurities 
is equal to the elimination of practically one and two-tenths tons of scale from the locomotive 
boilers each week, with the softener operating the full 24 hours. 



TABLE 11 

TYPICAL ANALYSES OP FEED WATER AT MAYPORT, FLA. 



No.1, 


No. 2, 


Untreated 


Water After 


Water. 


Treatment, 


Grains per GaL 


Grains per GaL 


18.09 


10.80 


8.57 


0.29 


6.88 


• • a • 


4.46 


1.08 


• • • • 


0.68 


0.70 


6.26^ 


2.48 


2.4S 


0.82 


• • * • 


0.88 


0.11 


18.89 


1.48* 


8.18 


9.82 



Total aolida 

Calcium earbraate . . . 

Caldum sulphate 

Masnesium carbonate 
Sotuum carbonate. . . 

Sodium sulphate 

Sodium chloride 

COi, free 

Iron and silieate 

Incnisting solids 

Non-^ncrusting solids . 




transformed into mono-carbonates. It is impossible to remove* chemically, sU the calcium and magneaittm as the memo- 
carbonates are sligbtly soluble. 



CHAPTER X 

STRAM £N6m£S 

Medianism of the Redprocating Engine. The working parts of a simple slide-valve engine 
are shown by Fig. 1. 

Steam from the boiler is piped to the steam chest C and admitted to the cylinder through the 
steam ports P' and P. The driving force of the steam is conmimiicated to the engine crank 
through the piston, piston rod, crosshead, connecting rod and crank pin. 

While the piston moves from one end of the cylinder to the other the crank shaft turns 
through one-half of a revolution. The piston makes two strokes, one forward and one return, 
for each revolution of the crank shaft. The distributicm of the steam to the cylinder is accom- 
plished by means of the valve D, The action of the valve is to admit steam alternately to each 
end of the cylinder and on the opposite stroke to allow the expanded steam to escape through 
the exhaust port J. 

Automatic Cutroff Vabe Gear (Fig. 2). The valve D alternately imcovers the steam ports 
P and P, when the piston has reached the end of its stroke, and allows live steam to flow into 
the cylinder driving the piston toward the opposite end. 

The valve receives its motion from the eccentric E located on the crank shaft, the motion 
of the eccentric being oommimicated to the valve by means of the eccentric rod R and valve 
stem H. The following description of the valve action and steam distribution refers to the 
head end of the cylinder. 

The piston / is shown at the '' head end " of the cylinder having nearly reached the end 
of its travel to the left. At this time the valve D is moving to the right and, for the position 
of the pi8t<m as shown, is ready to imcover the steam port P' and admit live steam back of the 
piston. 

, The relative positions of the crank pin and eccentric center, at this time, are shown by the 
diagram A as 1 and 1'. 

The piston continues to move to the left until the end of the stroke is reached while the 
valve continues to move to the right, and when the piston has reached the end of its travel the 
valve has partially imcovered the port P\ The object of opening the steam port slightly before 
the piston has completed its stroke is fw the purpose of preventing, as far as possible, a drop 
in pressure at the beginning of the stroke due to the throttling action of the steam in passing 
through a restricted opening. 

When the piston has moved toward the '' crank end '' to the line marked " cut off " the 
relative positions of the crank pin and eccentric center are indicated as 2 and 2' on the diagram A, 
During this portion of the piston stroke the motion of the valve has been reversed, owing 
to the angular advance of the eccentric center, and the valve returned to the position shown cut- 
ting off the steam supply. The steam port remains closed (valve moving to the left) until the 
piston has reached the point marked ** release " in its travel to the right. 

At ihia time the inside edge of valve D has reached the inside edge of the steam port P', as 
shown on diagram C, and the expanded steam begins to flow from the cylinder into the ex- 
haust port / which communicates with the atmosphere, a condenser or a low-pressure heating 
sjBbem. The relative positions of the crank and eccentric are 3 and 3' (diagram A), The 

229 



POWER PLANTS AND REFRIGERATION 



STEAM ENGINES 2S1 

piston, haviiig reached the end of its stroke to the right, has its motion reversed; the motion of 
the valve, however, is not reversed mitil a somewhat later period. . The expanded steam is forced 
out of the cylinder mitil the piston has reached a point in the return stroke marked " oom- 
presion," the valve having returned to the position as shown by diagram C. From this point 
on, until the piston has reached "admission," the exhaust steam remaining in the cylinder is 
oraipreesed in the clearance space. 

Tlie vahre action and steam distribution f ot the crank end of the cylinder are the same as 
described for the head end. r 

Tlie vahre action is most conveniently studied by means of a pasteboard model of the valve, 
crank and eccentric. The relative positions of the valve and piston for various parts of the 
revoluticm are readily determined by projecting vertical lines through the crank pin and eccentric 
outers as indicated. 

The valve is said to be in " mid-position " when it has reached the middle of its travel as 
shown on diagram E. The distance the valve extends over the outside edge of the port P' or 
P when in mid-position is termed the steam lap. The distance the valve extends beyond the 
inside edge of t^e port is termed the exhauti lap. 

Valve Design. In order to determine the dimensions of a slide valve, for any desired cut 
dSi and steam-port opening, the following method may be employed. 

The width of the ports P* and P are determined from a consideration of the area required 
to pass the volume of steam necessary without excessive pressure loss. 

Let 8 — average piston speed ft. per min. 
» 2 X r.pjn. X stroke in ft. 
A ss area of piston, sq. ft. 
a — area pori, sq. ft. 
V » average allowable velocity of steam through the steam port (5500 ft. per min. 

i^proximately) . 
I ss length of port, feet (0.82 X diam. cylinder approximately). 
w » width of port, feet. (To be not less than yi" to ^" for good castings.) 
a = Z X w. 

aV '^ AS ^ 

AXS 



V) 



0.82D X V 



Draw the crank pin circle to any convenient scale, and locate the position of the crank pin 
tor " admiasiCTi " and " cut off " (points 1 and 2). It is evident from an inspection of the dia- 
gram that while the crank pin was moving from 1 to 2 the valve must have opened and closed 
the steam port on the head end of the cylinder by an amount equal to the port opening desired. 

The center of the eccentric travels through the same angle (fi) as the crank pin. 

Lay <^ the angle fi, diagram B, and bisect it with a horizontal line. Then with O as a center, 
find by trial a radius (r), which will be the eccentricity of the eccentric, such that the horizontal 
dktance betwem the chord x ^ x and the circle is equal to the port opening required. Transfer 
to diagram A as shown. 

The total angular advance a of the eccentric, valve travel and steam lap are now known. 

Locate position 3 of the crank pin and eccentric center 3' for " release." Draw a vertical 
line y — y through eccentric center for release. The exhaust lap required is the horizontal 
distance between y -^ y and the center line of crank shaft. The line y-^y cuts the path of 
the eccentric center at 4' which is the position of the eccentric center for " compression " from 
which the corresponding position 4 of the crank pin for " compression " is found. By the same 
method of proo^lure the steam and exhaust Is^ are determined for the opposite end of the 
cylinder using the valve travel already determined. 



POWER PLANTS AMD HEFRIGERATION 



Fta. 2. Stbaii Bhoinx with Automatic Cot Orr Omui, 



STEAM ENGINES 233 

The total length of the Talve tace u equal to 2 (aieam lap + width of port) + distance 
between the inside edges of ports- Having determined the dimenaiona of the valve and ecoen- 
tiic Ks indio&ted, they may be checked by means of a model. 

The following data may be employed for the deaign of a slide valve. 
Maximum cut off 0.625 stroke 
AdmissioD 0.9S stroke 

Release 0.95 stroke 

r the pOTt at manmum cut off and preferably for a cut off of 



GOVERNING MECHANISM 

A steam en^e is deeded to cany a certain prc-determined " normal load " which is ordi- 
nsiily about 70 per cent of the maximum power it will develop with the same initial 
[vesBure and speed. This is equivalent to saying that tiie eafpiie will cany a fifty pa cent 
ovo'load. The office of the governing 
mechanism of an engine is to regulate or 
diange tite power input to meet the vary- 
ing demand of power output or load. It 
is f urthermoie essoiUal, especially in eleo- 
trical woik,- that the rotative speed re- 
main practically constant for a range of 
load from " no load " to " full load." 

The voltage generated by K «lynamo 
is prtqiortioiial to its rotative speed so 
that a chaitge in ^leed means a change 
in voltage. Incandescent lamps and 
motors are designed to operate, for highest 
efficiency, at some particular voltage bo 
that any change in the line voltage is im- 
desirable. 

The govnning mechanism is there- 
tan called upon to preserve a balance 
between the impellmg and resisting efforts 
which is the essential condition for uni- 
form rotative speed. 

Two methods for automatically ob- 
taining tha above results are employed 
for reciprocating steam engines. The first 
method is known as throUU goimrmng 
and is accomplished by varying the pres- 
sure of the steam supply. The second 
method is known ee eui <^ gooeming 
and is accomplished by varying the 
amount of steam admitted to the cylinder 
by changing the " cut off." Fto. 3. TBBorrLB oovEBtfoa. 

Throttlft Governing. This method of 
mgine governing is not employed, in this country, on reciprocating engines requiring close 
regulation, its principal use being limited to steam-driven air-compresaois, pumps, fans, blow- 
ers and steam turbines. 



234 POWER PLANTS AND REFRIGERATION 

The construction of a t3rpical throttling type of governor is shown by Fig. 3. The balanced 
valve B controls the admission of steam from the pipe to the steam chest. As the valve is 
lowered it diminishes the free area (throttles) for the passage of steam into the chest, thereby 
causing a reduction of pressure and actuating force on the engine piston. 

The valve B is attached to a stem D; which is forced upward by means of the spring P, 
The governor weights O are revolved by the mitre gears M and hollow shaft H^ the gears reoerv- 
ing motion from the pulley P which is belted to the engine shaft. The action of the spring P 
is such as to just balance the centrifugal force developed by the fly balls acting throu^ the 
levers and spindle S at the safe normal speed of the engine. 

As the engine speed increases, due to a reduction in the engine load, the increased centrif- 
ugal force overcomes the spring resistance and causes the balls to take a position further out 
from the center of rotation, and in doing so moves the stem and valve downward. This outward 
movement of the balls is transmitted to the valve stem D by means of the bell cranks L, 

With an increase in load on the engine the speed tends to decrease, in which event the firing 
action, aided by gravity, moves the weights O in and increases the valve opening, admitting 
steam at a higher pressure into the chest. 

Any increase in the engine speed above normal will therefore reduce the pressure and supply 
to the engine, and a decrease below normal will increase the pressure and supply. A balance 
between the impelling and resisting efforts is thus maintained, and with a properly designed gov- 
ernor the speed will remain fairly constant within limits. 

This type of governor is ordinarily employed on plain slide-valve engines having a fixed 
cut off at approximately ^ of the stroke. 

Cut Off Governing. An automatic cut off engine is rated to deliver a certain normal 
output with a cut off ^ to H of the stroke. 

If the power demand is less than normal, cut off automatically occurs earlier in the stroke, 
and if above normal occurs later in the stroke. 

The volimie of steam admitted to the cylinder, at constant pressure in this case, determines 
the amount of power developed in the steam cylinder. 

The variation in cut off is accomplished by means of the swinging eccentric (Fig. 2). 
The position of the eccentric center in reference to the center of the crank shaft is regulated by 
the shaft governor. The full lines j indicate the position of the eccentric, governor weight and 
arm for ^i cut off. When the load on the engine becomes less than normal, the speed immediately 
increases; the increase in the centrifugal force acting on the governor weight overcomes the 
spring tension, causing the weight to swing outward. 

This motion is transmitted to the eccentric through the pivoted weight lever and link, caus- 
ing the eccentric center to move toward the shaft center. The eccentricity of the eccentric is 
thereby reduced, resulting in a shortening of the valve trav^ and an earlier cut off. With only 
the friction load on the engine, the speed is a maximum, the governor weight taking the extreme 
outside position k. The eccentricity of the eccentric and valve travel are now a minimum 
and the valve is only moved a distance sufficient to uncover the port a very slight amount; 
simply enough to keep the engine running at the desired speed. 

In case the load on the engine becomes greater than normal, the speed decreases with a cat' 
responding decrease in the centrifugal force acting on the governor weight. The spring tension 
being greater than the centrifugal force at the reduced speed pulls the weight toward the shaft, 
and at maximum load the speed is a minimum, the governor weight taking the position i. Tlie 
eccentricity and valve travel are now a maximum, giving the longest cut off obtainable with the 
gear. Fig. 4 shows a conunercial form of this type of shaft governor. 

The maximum cut off obtainable for this type of valve gear, as usually designed, is i^proxi- 
mately ^ of the stroke, the maximum overload capacity being approximately 50 per cent on 
a basis of 14 cut off for normal load. Engines equipped with shaft governors of the type shown 
may be made to regulate within two to four revolutions per minute from no load to ftdl load. 

In practice this type of governor is frequently equipped with a double set of weight arms 



STEAM ENGINES 



ring ft symmetricaJ form with refwenoe to the crank ohaft. Thia aoi>Btnioti<m 
b to obviftte the poHtbOity of oacillMtioa, due to the govcntor not being in a gravity balaiiaod 




Pio. 5. Inkkiia Trra Soait Govbiinor, Pobitioh Showh ros MAxmDu om orr. 



The In«rlia Goranor. The govemor previoualy deacribed depends idnioat entirely o 
BctioD of centrifugal force for its operation and is the original type aS shaft goveraor. 

The praaent type of shaft govemw (Fig. 5) employed by the majwity (rf hi^ speed ex 



236 POWER PLANTS AND REFRIGERA1I0N 

builders is known as the RiUs inertia governor. With this type/ in addition to the oeoftrifugml 
force acting on the weights, the inertia of the weights aids in the rQgiilati<ML 

When the engine tends to speed up, the inertia of the weights causes them to lag and aid the 
spring in reducing the throw of the eccentric and shortening the cut off. 

A decrease in the engine speed will, of course, produce the opposite effect. 

This type of governor is particuliu-ly sensitive to a change oi speed, and therefore a doae 
regulator, a regulation of 1^^ per cent being common for a gradual change of load from no load 
U>full load. Fig. 6 shows a high speed engine equipped with the RUe9 inertia governor. 

Coefficient of Regulation. The coefficient of speed regulation or sensitiveness of a govern- 
ing mechanism, as implied to steam engine practice, is defined as the percentage of variatioa 
oi rotative speed from the full load speed. 

Let S B the mean speed of the engine at normal or full load in rev. per min. as specified 

by the builder or as determined by test. 

Si » the maximum speed obtained (no load speed). 

V » the coefficient of regulation or sensitiveness specified in per cent. 

VS 

— - » aUowable change of speed above normal in rev. per min. 

Then Si must not exceed S -h — . 

100 

Example. The speed of an engine as specified by the builder for full load is 196 r.p.m. and 
the governor is guaranteed to regulate within 1.5% for a slow change in load from *' no load ** to 
** full load," or vice versa. 

The speed of the engine was actually 198.8 r.p.m. when operating under '* no load " (friction 
load) conditions and for *' full load ** the speed was found to be 196 r.p.m. 

The speed variation is 

VS 100 X 2.8 

— ■= 198.8 — 196 or 2.8 r.p.m. .*. V — — — — or 1.4%. Th^ engine, therefore, fulfilled 

the guarantee in this respect. 

Unless otherwise noted, it is assumed that the change from " no load " to " full load " or 
vice versa is made gradually and not suddenly. Frequently two regulation guarantees are 
made: one for a gradual change of load and a separate one for a sudden change of load from 
" full load " to '' no load,'' the latter being the greater of the two. 

IxL order to determihe the change in speed due to a sudden variation of load when the engine 
is direct-connected to a generator, the full load is first applied; then the main switch is opened. 
The speeds are determined by means of an electrical tachometer, which is simply a small generator 
in circuit with a milli-voltmeter; the milli-voltmeter having been previously calibrated to read 
the speed direct. 

Bzample. Assume that a speed of S ^ 196 r.p.m. is obtained when the machine is operated at 
full or normal load, and when the load is suddenly removed the maximum instantaneous speed recorded 
IB Si ^ 200 r.p.m. Tho maximum variation is: 

— - 200 - 196 = 4 r.p.m. .*. V = 2%. 

Specificationa ahotdd clearly state the method to be employed in determining the speed variation 
and basis upon which the calculations are to be made. This is particularly important when the 
unit is supplying both a lighting and rapidly fluctuating motor load, as in this case the instan- 
taneous variation of speed must be limited to a small margin to prevent '' blinking" of the lights. 

For high-speed direct-connected units the U, S, Treasury Department specifies that the maxi- 
mum variation in speed for a slow change in speed from "no load '* to "full load " or vice versa 
shall not exceed 1 14 per cent of the speed at fidl or normal load, and that for a sudden change in 
load the maximum variation shall not exceed 2 per cent. 



STEAM ENGINES 237 

Tba Balaiic«d Dotibl»-Portod Valre. (Fifp. 7 and 8.) The shaft governor is held in equi- 
libriuiQ under the opposing forces of spring tension And centrifugal tixce. Any externally applied 



force, such as the power required ta move the valve and the force required to overcome the 
inertia or accelerate the reciprocating parts of the valve gear, tends to disturb this 



238 POWER PLANTS AND REFRIGERATION 

The force required to more an unbalanced elide ralve u conaiderabk^ being the product of 

the projected area of the vahre, steam preesure in the chest and the coefficient (rf friction <^ a 

lubricated surface. 

The accelerating force 
required to reverse the mo- 
tion of the valve and gear 
is a functifm of the weight 
and linear vtitxity of the 
reciprocating parts of the 
gear. 

In <vder to minimise 
the effect oi friction die 
valve is provided with a 
cover plate, thus preventing 
the steam pressure from 
acting on the top side. A 
valve thus equipped is 
known as a balanced ralve. 
In order to reduce the 
effect at tbe inoiia taroea, 
the Imgtli of travel is re- 
duced by miJfing the valve 
Fio. 7. DouBLB-PoBTXD Baumced Valvb. doubft-porUd. With this 

arrangement the same area 

of port opening is obtained by one-half the movement, other comUtions being equal; the steam 

lap is only one-half as large as tor the plain slide valve. 

The Corliss Enghie. The essential fea- ^^^ 

ture of the Corliss type of eteam engine is the 

employment of four cylindrical valves; the 

upper two, with a horiiontal engine, being used 

for steam admission and the lower pair for ex* 

hauat. The principal parts of the Ccahm cyl- 
inder and valve gear are shown by Figs. 9 to 15. 
The object of the Ccwliss engine, which is 

obvioualy more expensive to build than the 

shde-vdve type, is a reduction in the heat loss 

incident to wire drawing of the steam dunng 

admission, initial CMujensation and the detri- 

mental eEFect of a large clearance volume. 

With the slide-valve engine the live steam 

must travel through the same port as the 

eadiaust steam, the result being a cooling of 

the ateam admitted and a greater initial con- 
densation than occurs when separate steam 

and exhaust ports are provided. The steam 

porta ot a slide-valve engine are gradually extmt 

opened and closed, resulting in a loss in prea- p,^,. g, g,cnoM Thbouoh Bioa-ann. BiraiNa 

sure incurred in forcing the steam through a CrUNDia. 

restricted port opening. 

With the Corliss engine ttus loss is reduced to a minimum, dus to the fact that the steam 

valvee are opened rapidly by the valve gear and closed promptly by the action of dash pots 

operating independently of the gear. 



STEAM ENGINES 239 

Experience has demonstrated that an engine birring a lai^ clearance volume (spaoe be- 
tween the piston and valve covering the port when the piston ia at the end of its stioke) is not 
so eccmomical in the use of steam as one with a smaller clearance. With Corliss valvea the 




Fig. O. VssncAii Swtiiok Thbouoh a Ooauss Otumpbr Showtho Donsu-PORrSD Valww 




dearoQce volume is several times smaller than is possible with a single slide valve. The net 
Rsult is aa oigine with an increased steam economy. 

Use valves are operated as indicated by the diagram of connections (Fig. 10). The motion 
of the eccoitoio is transmitted by the eccentric rod and hook rod to the wrist plate. The oe- 



240 POWER PLANTS AND REFRIGERATION 

ciUatory motion of the wrist plate is transmitted to the valves through the steam and exhaust 
rods, rockers and valve stems. 

The exhaust valves are at all times connected with the wrist plate, "whereas the steam valves 
only receive motion from the wrist plate during the period when they are opening the steam 
ports for admission of steam to the cylinder. 

The diagram shows a single eccentric gear; that is, both steam and exhaust valves driven 
by the same eccentric. The gear is shown in its central position, the crank pin indicated by 1 
and eccentric center by 1'. The crank pin is in '' lead '' position, the steam valve for the head 
end of cylinder being just ready to open or already shghtly open. 

The maximum cut ofif must occur when the center of the eccentric has reached the end of 
its travel or at 3' (corresponding position of the crank pin at 3) if the valve is to be closed by the 
action of the dash pot, as the steam rocker (Fig. 13) is at the highest point of its travel at this 
time. .* 

The exhaust valve uffst close (compression) at 4 somewhat before admissi<Mi (1) ocoara 
and will open (release) when the center of the eccentric is at 5', found by drawing a perpendicular 
through 4'. The total angular advance then determines the position 5 of the crank pin for release. 
To lay out the Corliss gear a ''cut and try" method is adopted. 

In order for release to take place somewhat before the piston has reached the end of its travel 
and at the same time obtain some compression, it is foimd with the single eccentric gear that 
the TnayiTniiTn cut off obtainable is approximately 3/8 of the stroke. 

With two eccentrics, one for the steam and one for the exhaust valves, a cut off of i^proxi- 
mately 7/10 of the stroke is obtainable. 

The valve stem F, for the steam valve for the. head end of the cylind^, is keyed to the 
steam arm D (Fig. 11), on the end of which is bolted a hardened steel plate. The "steam arm" 
is lifted by the '' steam hook " or latch L, which is also provided with a hardened steel plate, 
which engages with the steam arm as shown. 

Lifting the steam arm opens the steam port as indicated by Fig. 10. The steam hook is 
attached to one arm of tlie '' steam rocker '' A, the other arm oi which is connected with the 
wrist plate by means of the steam rod. * The steam rocker has a bearing on the steam valve 
bonnet through which the valve stem V passes. Fig. 13 shows an assembly of the releamng 
gear. 

When the piston has reached the end of the stroke, on which the inlet gear shown is located, 
the wrist plate is pulling the rocker which in turn, through the steam hook, is raising the steam 
arm and opening the steam port. ^ When the piston has reached the point in its travel wh»e 
cut off is to take place the arm B of the steam hook (Fig. 12) engages the " knock-off ** cam C and 
the hook L is disengaged from the steam arm D, The steam arm D is attached to the drop 
rod, the other end of which is connected to a vacutmi cylinder termed a " dash pot.'' (Fig.l4.) 
When the steam arm is raised a partial vacumn is created in the dash pot so that when released 
by the steam hook, the steam valve is rapidly closed. The piston in the dash pot is so arranged 
that a small quantity of air is trapped on the down stroke which acts as a cushion and {E^vents 
the piston or plimger from striking the bottom of the air cylinder. 

The knock-off cam C is attached to the cam lever F which has a bearing on the steam bonnet. 
The cam lever F is operated by the governor as* indicated by Fig. 15. When the engine i^)eed 
tends to increase above normal, due to a decrease in the load, the governor weights move out- 
ward. This motion is commimicated to the sliding collar, by means of the lifting Unks, to which 
is attached the governor drop rod connected with the bell crank. Raising the collar moves the 
knock-off cam C (Fig. 11) in a downward direction so that the hook arm L will engage the cam C 
earlier in the stroke, thus producing an earlier cut-off. An increase in load results in a decrease 
in speed and the action described is reversed, causing the cut off to occiu* later in the stroke. 

In case the belt driving the governor should break, the ** safety cam " ^ is thrown into action 
and prevents the latch from engaging the steam arm. 

In order to prevent the operation of the safety cam, when shutting down the engine, a pro- 



STEAM ENGINES 





Fia. 11. Ttpicai. RELEAsnca Oi 



Flo. 12. BxTBivE PounoN or Hook. 



Tio. la. AmmitBi.T OF OoBLoe Bii^kasino Qmab. 



Vm. 14. Dash Per. 



242 POWER PLANTS AND REFRIGERATION 

Jeoticm on Ute collar oomes to rest on a vertical arm (governor safety atop) either manually oper- 
ated M' BUtomAtieally operated by the steam pressure below the engine throttle valve. 

With a single eccentric releaainf; type of gear the m&jdmum cut off obtainable ia ^proxi- 
mately li of the stroke, which limits the overload capacity of the engine to about 25 per cent 
on a basis of H cut c^ for the normal load. 

In ardo' to increaae the range lA cut off to approximately seventy per cmt of the stroke 

... ,.>•• 



y 



<?."' 



,• 



..•^ 



Fra. 16. OoraBNiwi Micbanisu or tbe Cosuaa BiraiNn. 

a separate eccentric is employed to opa«t« the exhaust valves. The msj<nity o! Corliaa engioea 
■re BO equipped in order to cany a fifty per cent overload. 

The speed of an engine equipped with the Corliss type of gear ia limited in practice to ap- 
proximately 125 r.p.m. as at higher speeds the hook, as ordinarily constructed, does not always 
latch or engage the steam arm. With gears especially designed, speeds of 100 r.p.m. have be^i 
obtaioed. 

STEAM ENGINE PROPORTIONS 

In the design oi steam en^nes, the pressure on the wearing surfaces must be taken into 
oonsideration when proportioning the various details. The pressure on the several parts is due 
not only to weight but to the motion of the reciprocating parts acting under the steam pressure. 
Fot example, stesm pressure chi the piston is transmitted to the crank pin tlirough the pistwi 
fod and connecting rod, which may be resolved into two components. 



STEAM ENGINES 243 

One of tbe oomponents, which acts radially along the crank, exerts a pressure on the main 
bearings which is in addition to the {Nressure caused by the weight of the shaft, crank, flywheel, 
etc The other component acts at ri^t angles to the crank exerting a tangential piessure, caus- 
ing the crank to revolve and is known as the crank effort. Croes-hcAd pins are also subjected to 
pressure from piston and connecting rod. 

FrcMn experience based on practical design and ccxistruction, permissible pressures per square 
indi OQ the areas obtained by multiplying the length by the diameter oi journal have been de- 
duced, which are as foUows: Main bearings, 140 to 160 lb.; crank pins, 1000 to 1200 lb.; cross- 
head pins, 1200 to 1600 lb. 

Based on these values and with steam pressures oi 100 to 125 lb., the following data by 
James B, Sianwood have been compiled as conforming to common stationary engine practice and 
will be found useful in checking the dimensions submitted in an engine proposaL It is not within 
the province of the text to treat on the mechanics oi the steam engine. The problems arising in 
the design of governors and the determination of the weight of flywheel, when the engine is re- 
quired for parallel operation of alternating-current machines, are very complicated. 

Relatton of Engine Parts to Piston 

Relation to 

Piston Piiton 

Diam. Area 

Main Shaft, diameter 0.42—0.50 

Main Shaft, length 0.85—1.00 

Crank Pin, diameter 0.22—0.27 

Crank Pin, length 0.25—0.30 

Croeshead Pin, diameter 0.18—0.20 

(>oediead Pin, length 0.25—0.30 

Piston Rod, diameter 0.14-0.17 

Steam Port Area: 

Slide Valve Engine 0.08—0.09 

Hi^ Speed Auto. Engine 0.10— 0. 12 

Coriiai Engine 0.07—0.08 

Exhaust Port Area: 

Slkle Valve Engine 0. 15— 0.20 

Hi^ Speed Auto. Engine 0.18—0.22 

CorliflB Engine 0. 10— 0. 12 

Steam Pipe Diameter: 

SUde Valve Engine 0.25-0.25 + H" 

Hi^ Speed Auto. Engine 0.33 

Coriias Engine 0.30 

Exhaust Pipe Diameter: 

Slide Valve Engine 0.33 

Hi^ Speed Auto. Engine 0.375 

CorilsB Engine 0.33—0.375 

Per Cent 
Piston Displacement 
Qearanoe Space: 

Slide Valve Engine 6—8 

High Speed Auto. Eng., 1 valve 8—15 

High Speed Auto. Eng., 2 valves 3—5 

Corliss Engine 



244 POWER PLANTS AND REFRIGERATION 

Lb. per rated 
Horsepower 
Weights of Engines: 

Slide Valve. , 126 — 135 

H. S. Auto 90 — 120 

Corliss ; 220—250 

Weights of Flywheds: 

Slide Valve Engines 33 

High Speed Auto. Engine 25—33 

Corliss 80—120 

Rules for Flywheel Weights, Single-Cylinder Steam Engines 

Let d s diameter of cylinder in inches; ^ 

S — stroke of cylinder in inches; 

D » diameter of flywheel in feet; 

R » revolutions per minute; 

W B weight of flywheel in pounds. 

For slide valve engines, ordinary duty, 

IT = 350,000^^ 
For slide valve engines, electric lighting, 

ir- 700,000 ^P^ 

For automatic high-speed engines, 

W = 1,000,000 



For Corliss engines, ordinary duty, 

d^S 
W - 700,000^^ 

For Corliss engines, electric lighting, 

d!^S 
W = 1,000,000 



THE WORK DIAGRAM 

The action of the steam in the cylinder of a reciprocating engine is conveniently studied 
by means of a woric diagram (Fig. 16). 

Assume that steam at an initial absolute pressure of Pi lb. per sq. in. is admitted to 
the cylinder on the left-hand side of the piston for the entire length of the piston stroke Iq. 
This is the condition of operation of a direct-acting steam pump. The absolute pressure Pij ex- 
isting on the opposite side of the piston, termed back pressure for an engine exhausting te 
the atmosphere, is approximately 1.5 lb. per sq. in. above the barometric pressure. This small 
increase is necessary to overcome the friction of the steam in its passage through the port pas-, 
sages and exhaust piping. 

The effective driving pressure is therefore Pi — P% lb. per sq. in. 
Let A ^ area of piston in sq. in. 

Li = length of piston stroke in ft. 
W ■= work per stroke ft.-lb. 
- (Pi - P,) UA 



STEAM ENGINES 



245 



Ft "■ piston displaoement Tohune cu. ft. 

/. TT - (Pi - P,) Vi 

The work W is represented by the area CDHO of the diagram shown to the scale used for 
prasure and volume. 

C D 

/> //=^ J^=tf constant 



Pressure 




Fio. 16. Thk Work Diagram. 

Let the length of the cylinder be increased, as shown by the dash lines, so that the length 
of piston stroke is now Lt ft. and the piston displacement volume Vt ^ LtA cu. ft. 

If the same volume (Fi) at pressure Pi be admitted to the lengthened cylinder and the 
supply cut off when this volume has been admitted the steam will expand with a decreasing pres- 
sure as the pistcm moves to the ri^t and the volume increases. The relation existing at any 
point of the stroke between the pressure and volume, during the expansion period, is ap- 
proximately given by Boyle*$ law, viz., pressure times volume is a constant. (PV « a constant.) 

Let Pfl a- the terminal pressure at the end of the stroke. 



Then PiFi ^ PiFt or Pi 



PiF 



Pi Vt 



- and — = — 
Vt P. Vi 



The ratio jr or the number of times the steam is expanded is termed " the ratio of expan* 
Vi 



•wfi" (r). 



Vi Lx 
Am the area of the piston is a constant, we have the relation tt '^ y* The recfp- 

Vt Lt 



L 



246 



POWER PLANTS AND REFRIGERATION 



rocal of the number of expanaioiifl f — j is the fraction oi the stroke completed when the aapply 

1 La 
of steam vnm cut off from the cylinder. The ratio — or -=- is termed the " cut off" >• 

r Lt 

Thus with an initial absolute pressure, Pi « 100 lb. per sq. in., Vi «- 1 cu. ft. and Ft >- 

4 cu. ft., the terminal pressure P| » 100 X }i * 25 lb. The ratio of expansion is r >-•—-*■ 4 

Vi 



Admtiiton 



\Catafr 




Bxhsutt 
Fig. 17. Mbtbod of Oonbtbuctinq Bzpambion Cubtb 



and the cut off is said to have taken place at }^ <»> 25 per cent of the stroke. 

It is evident that the terminal pressure Ps can never be less than Pt if woric is to be done 
on the piston by the steam throughout the st^t^e, which is the effect naturally sou^t. 

The limit of the terminal inressure is therefore fixed )>y ^Ps^or when Pt •■ Pt. 

In practice, the engine cylinder is designed for a somewhat less number of expansions as 
will be later explained. 

The expansion curve DE may be plotted by pwats as may be determined by calculating the 

P V 

pressure as P« f ot a length of stroke L, corresponding to volume V„ thus P« «- — — . 

The expansion curve is a h3rpeibola and may be constructed graphically as diown by Fig. 17. 
The average fcnrward pressure during the expansion part of Uie stnke, from D to E or H 
to F, may be i4>proximately determined by dividing the length HF (Fig. 16) into a number (n) of 
equal parts and scaling the height (x) of the mean pressure ordillate for each division. The sum 
<^ these ordinates (xi + Xt + ^ • • • ^) divided by n and multiplied by the vertical scale 
of pressures gives the mean abdoltUe forward prewure P during the e]q)ansion period. A more 
exact result is obtained by integrating the area DEBL by means of a planimeter and dividing this 



STEAM ENGINES 247 

are* by the length LB and multiplying the quotient (height of mecm wdinate) by the vertical 
preasure scale. The mean effective pressure acting on the piston during (A« expantion portion 
tftim Uroke is therrf«o P - P. lb. per sq. in. The wwt tr* performed ia (P - P.) (V, -V,) 
ft.^. Ilie total m^ for the entire stroke is ther^we. 

W "Wi + W' = (Pi- PO V,-ir{P- PO (Vt - V,) fL-lb. 

Econonv of Using Stesm Eqwiuively. Hu per cent gain in work obtained per unit tgJ- 

nme or ira^t at steam used, ficmi the foregoing, is tbenfore — ~ — r X 100. 

Tbe reason tcv using Bt«un eiqMnaively in an engine cylinder is thus apparent. 
Theoretical Heaa BflediTe Pte wur o. Hie conventional method for determiaing the th«- 



Fra. 18. SaonoN or SraAV BNom tsmotram. 
J mean eSective pressure (m.e.p.) for the entire piilon ttrdke when the uutial pressure (Pi), 

back piOBUre (P|) and ratio of ezpansioa [r— — 'j or cut off f — j are given or aasumed is 06 

foOowa: 

Area CDEBA - Area OFSA 
m.e.p. -P. w^k iR 



248 POWER PLANTS AND REFRIGERATION 

Area QFBA = PiVt, AB = Vt. 

Area CDEBA = Area CDLA + Area DEBL. 

Area CDLA = PiVi, The area DEBL under the expansion line by method of cal- 



AB 
cuius ^ DLX ALX hyperbolic log. —-:. DL = Pi. AL 

AL 

e3cpansion. 

/. Area DEBL = PiVi log, r. 

PiFi + PiFi log. r - P,7, 



AB V, 
Vi. -r- = — - r, the ratio 
AL Vi 



p. = 



F, 



The value of the expression 



= -^(H-Iog.r)-P..(---J 
.P.(L±^)_P. 

1 + lofe r 



may be taken from ' Table 1 for various as- 



sumed ratios of expansion (r). This is the ratio of the mean forward pressure to the abeolute 
initial pressure. 

Example. Required the theoretical m.e.p. for the following conditions. Initial pressure 100 
lb. per sq. in. gage. Pi -» 100 + 14.7 " 114.7. Back pressure 2 lb. gage Pt» 16.7. Ratio of ex- 
pansion r B 4 (cut ofiF yi stroke). 

From Table 1, ^.±i2?f_r « 0.6965. 

r 

Theoretical m.e.p. - (114.7 X 0.5965) - 16.7 - 51.7 lb. per sq. in. 

TABLE 1 



Ratio of 


Cut-off 

Vr 


Ratio of Mean 


Ratio of 


Cut-off 

Vr 


Ratio of Mean 


Expansion 


Forward to Ini- 


Expansion 




r 


tial PreMure 


r 


tial IYmbui e 


SO 


0.033 


0.1467 


6.00 


0.166 


0.4668 


28 


.036 


.1547 


5.71 


.175 


.4807 


26 


.038 


.1688 


5.00 


.200 


.6218 


24 


.042 


.1741 


4.44 


.226 


.5608 


22 


.045 


.1860 


4.00 


.250 


.5965 


20 


.050 


.1998 


8.63 


.276 


.6808 


18 


.056 


.2161 


3.33 


.300 


.6615 


16 


.062 


.2358 


8.00 


.838 


.6995 


15 


.066 


.2472 


2.86 


.350 


.7171 


14 


.071 


.2699 


2.66 


.376 


.7440 


18.33 


.075 


.2690 


2.50 


.400 


.7664 


13 


.077 


.2742 


2.22 


.450 


.8095 


12 


.083 


.2904 


2.00 


.600 


.8466 


11 


.091 


.8089 


1.82 


.650 


.8786 


10 


.100 


.3303 


1.66 


.600 


.9066 


9 


.111 


.3562 


1.60 


.625 


.9187 


8 


.125 


.3849' 


1.54 


.650 


.9292 


7 


.143 


.4210 


1.48 


.675 


.9405 


6.66 


.150 


.4347 


• • • • 







Actual Mean Effective Pressure* The actual mean effective pressure of a reciprocating 
engine is determined by making use of an instrument termed an indicate. An instrument of 
this kind is shown diagrammatically by Fig. 17 and the actual instrument, in section, by Fig. 18. 

The indicator consists of a small cylinder in which a piston operates. The pisUm is at- 



STEAM ENGINES 



249 



tadied to one end of a helical spring, the other end of the spring being attached to the indicator 
Qrlinder. Hie indicator cylinder is placed in communication with the engine cylinder by means 
of the pipe nu The pressure of the steam on the piston (orcea it upward against the resistance of 
the firing. The distance the piston rises is directly proportional to the steam pressure. 

The motion of the piston is transmitted to a pencil n through the medium of a combination 
of levers (parallel motion) such that the vertical motion of the pencil is exactly parallel to the 
center line of the indicator cylinder. 



TABLE 2 

OOBCPARISON OF MEAN EFFECTIVE PRESSURES OBTAINED IN PRACTICE WITH TABULAR 

AMOUNTS 

(NonrOmdMising Ihigines) 



Type of EngliM 


Siaa, Inehes 

* 


Steam Pipe 
Pressure 
Lb. Gage 


Actual 
Cutoff 


Actual 

from 

Diagrams. 

Lb.per Sq.In. 


BA.E.P* 
as per 

Usual 

MTr'a 

Tablea 


SfaiilaTalY* 


8 xl2 


88 

82.4 
81.9 


0.208 
.812 
.878 


' 24.6 
84.8 
48.1 


86 
48 




65 


Sfaiila tbIt* 


18 X 12 


100 


.82 
.42 
.68 


46.6 
66.6 
67.6 


64 

76 




84 


Oidim 


28 z60 


72.8 


.367 


88.1 


46 






Oidim 


28HX69H 


101 


.816 


41.2 


61 






Cflrito 


16 z42 


100 


.178 
.281 
.828 


29.6 
88.1 
48.4 


40 
49 




64 


Coritas 


22 z80 


148.6 


.201 


64.6 








Gridiftm tbIt* 


28 z60 


66.1 


.222 


28.8 


80 






rvu^Talvv 


19 zl8 


100 


.286 
.186 


42.8 
29.3 


68 
42 







The pressure, in pounds per square inch, required to move the pencil vertically one inch 
is termed the scale of the spring. Thus, if a 60-lb. spring is used and a pressure of 120 lb. per 
sq. in. existed within the cylinder, the pencil would move through a vertical distance of 2 in. 

The pencil presses against a piece of paper wn^pped around a drum. The drum is oscillated 
about its axis by means oi a chord, one end of which is wrapped around the lower part of the 
drum and the other end attached to a reducing motion, which receives its motion from the 
engine cross-head. The drum movement is therefore a reproduction of the engine piston move- 
ment to a reduced scale. 

Hie diagram traced by the pencil is a pressure volume (PV) diagram, as the pressure exist- 
ing at any and all points of the engine piston stroke is automatically recorded on the palter. 
This diagram is termed an indicalor diagram. 

Owing to the imperfections in the working of the actual engine the diagram obtained is 
similar to the diagram shown by the dash lines (Fig. 17), and somewhat less in area and conse- 
quently has a smaller m.e.p. than the theoretical diagram. 

There is a loss of pressure during admission of steam to the cylinder due to frictional resist- 
ance (^ the steam passing through the steam port and also from the condensation of a small 
portion (^ the steam when it is brought in contact with the cooler cylinder walls. 

The esdiaust is opened (release occurs) somewhat before the piston reaches the end of the 
9koke and doses (compression begins) before the piston has reached the end of the return or 



250 POWER PLANTS AND REFRIGERAHON 

exhaust strdce. After the exhaust is closed the steam trapped in the cylinder is compressed in the 
clearance space of the cylinder, the pressure rising approximately in acocNPdanoe with Boyle's law. 

Admission occurs at H near the end of the return stroke. 

Diagram Factor. In order to calculate the size of cylinder required to perform a certain 

amount of woric an estimate <rf the expected m.e.p. is necessary. The ratio ^r : — ' — 

is termed the diagram fader. thewetical m.e.p. 

The theoretical m.e.p. referred to, is calculated as previously given lot the square-cornered 
card, assuming the cylinder as having no clearance space, full initial pressure up to Uie point 
of cut off release at end of stroke and no compression, and initial pressure being that ahooe the 
engine throttle. The data given by the following table may be used in this connection. 

TABLE 3 

APPROXIMATE DIAGRAM FACTORS* 
Rfttio of Aetual M.E.P. to Theoretical M.E J*. 



Type of Engine Simple Compound 




Sinfle-TalTe hicfa epeed 0.80 > 0.70 

Ftour-valve high epeed .86 .76 

CotUm dow end medium epeed .90 .80 

* Baeed on initial preMure above the throttle Talve. 

Example. Required the expected m.e.p. for a simple high speed engine for the fc^owing conditioDs 
of operation: Initial pressure 100 lb. per sq. in. gage (pi ■- 100 + 14.7 » 114.7), back pressure 2 

lb. per sq. in. gage (ps - 2 + 14.7 « 16.7). Cut off f - j 0.25 stroke, r » 4. 

1 + \o%g r 

- 0.5065 from Table 1. 

r 

Theoretical m.e.p. - (114.7 X 0.5965) - 16.7 - 51.7 lb. per sq. in. 

Expected m.e.p. — Theor. m.e.p. X diagraTn factor — 51.7 X 0.80 « 41.36 lb. per sq. in. 

Fig. 19 shows a reproduction of a pair of indicator cards taken from a medium speed four- 
valve engine and the method used in calculating the diagram factor. The " boiler pressure" 
refers to the pressure above the engine throttle. 

Example. {8m Cvi, OppoHU Page,) 

Diameter of piston, 15 inches. Net area of piston, head end, 176.71 sq. in. 

Diameter of piston rod, 2^ inches. Area of piston rod, 4.66 sq. in. 

Net area of piston, crank end, 172.05 sq. in. 

Head End Crank End 

Average ordinate length , 1 .0772 1 .0772 

Multiply by scale of card 60 50 

Mean effective pressure 53.86 53.86 

P^ L A N 

Crank end 53.86 X 1.667 X 172.05 X 150 



33,000 
P, L A AT 

Headend 53.86 X 1.667 X 176.71 X 150 



70.22 



72.11 



33,000 
Combined indicated horsepower 142.33 

(1 + log, 4 \ 
7^~7 - 14.7 = 57.87 lb. 

Dtagram factor = = 0.93. 

57.87 



STEAM ENGINES 



A 



261 



U 



'S99Jd tfOfiog sqi/oi 




'9S9Jd femuj 'sqi oOh 



~^ 



252 



POWER PLANTS AND REFRIGERATION 



POWER DEVELOPED BY RECIPROCATING ENGINES 

The power developed by an engine is expressed in terms of either brake hortepower (d Jip.) 
<»> indicated horsepower (i.hp.) usually the latter. 

Brake Horsepower. The brake horsepower is the power delivered by the engine crank 
shaft as determined by means of a brake mounted on the flyidieel (Fig. 20). 
Let D » diameter of flywheel measured in feet, 
n » revolutions of engine per minute. 
P « resistance at circumference of wheel, lb. 
R s load on scales, lb. 
a <B length of brake arm, ft. 



PX— -- RXa. 



PD = 2Ba 



1 hcMTsepower •■ 33000 ft.-lb. per min. 

tD Pn 2TaRn 



Brake hors^)ower (d.hp.) 



33000 



33000 



For a discussion of various forms of brakes and absorption djrnamometers see Carpenter and 
DiediricKa* " Experimental Engineering." 




Pio. 20. Front Bbakb. 

Example. What is the brake horsepower developed by an engine having a 5' ~ O'' diameter 
(D) of wheel running 200 r.p.m. (n), length of brake arm (a) 6' — CK', weight on scales iR) 200 lb. 

2 X 3.14 X 6^ X 200 X 200 



d.hp. "" 



33000 



- 46.7. 



Indicated Horsepower. The indicated horsepower refers to the power developed in the 
steam cylinder as determined from the indicator card and is greater than the brake horsepower 
by an amoimt equal to the frictional losses in the engine. 

Let P« » mean effective pressure lb. per sq. in. 
A = area of piston, sq. in.* 



* No ded ucti on it made for tbe tree of tbe pieton rod on tbe crssk end in ptiliminsry eelenlations for ttie 
ejrUader. 



STEAM ENGINES 253 



L » loigth of stroke, in feet. 

N B number of woridng strokes per min. 

» 2 X r.p.m. for double-acting cylinder. 
S » average piston speed ft. per min. 



Uip. 



P^LAN ^ P. AS 
33000 33000 



Example. Required the indicated horsepower of a \2f' X 12^' engine operating under the following 
jcx>ndiUons: Mean effective pressure P, - 40, r.p.m. - 300, N - 600, A » 113.1 sq^ in., L » 1 ft.: 

40 X 1 X 113.1 X600 

l-ho ■■ « 82.2 • 

^ 33000 

Machine Efficiency. The efficiency of a machine, in general, is a measure of its economic 
performance under certain imposed conditions of operation. 

There has been established, by custom and usage, several so-called efficiency standards by 
means of which we may conveniently compare the performance of various t3rpes of boilers, heat, 
hydraulic and electric motors, etc. The power output is alioays less than the energy or power 
input due to the transformation of a portion of the original mechanical energy supplied into 
heat energy caused by the friction of the moving parts. As this transformation serves no useful 
purpose it is an economic loss in the sense that a portion of the energy supplied does not reap- 
pear in the final desired form. 

Wh^i the friction loss of a machine is mentioned it is customary to state it as a percentage 
of the power supplied or power input. It, however, is more common to express or convey the 
idea as to magnitude of the friction loss not in terms of power input but by means of a ratio 
termed mechanical or machine efficiency. 

The mechanuxU or machine efficiency of an engine, fan, piunp, etc., is defined as the ratio of 
the useful work performed or power output to the mechanical energy supplied or power input, 
both input and output being measured in foot pounds or the more commonly used unit, the 
horsepower. Thus the machine effi^yiency of a steam engine is the ratio 

brake horsepower (output) d.hp. 

or . 

indicated horsepower (input) i.hp. 

The machine ^idency of a centrifugal pmnp is the ratio 

water horsepower (output) w.hp. 

: or 

brake horsepower (input) d.hp. 

The machine efficiency of a reciprocating steam pump is the ratio 

water horsepower (output) w.hp. 

or -: 

indicated horsepower (input) i.hp. 

The machine efficiency of a fan is the ratio of 

air horsepower (output) a.hp. 

or 



brake horsepower (input) d.hp. 

The machine efficiency of an electric generator is the ratio 

electrical horsepower output e.hp. 
brake horsepower input d.hp. 

Hie maehine efficiency of an electric motor is the ratio 

brake horsepower output d.hp. 

or ' 

electrical horsepower input e.hp. 



254 POWER PLANTS AND BEFRIGERATION 

The ebotrical horsepower ntlag tor genentora is based on the electrical oatput at genenUior 
terminals. 

The hcHaepower mting of motora refers to the brake honepower output of the machines. 

The machine efficiency of stationary engines baaed on testa run at normal load vary fnMn 
86 to 95 per cent. For preliminary oalculatJons 0.92 may be sasumed as a fair avovge. As the 
frietional horsepower is fairly constant for all loads, the machine efficiency tor the underloads is 
necessarily much less. 

Generator efficiencies vary from 90 to 97 per cent at rated load, tending to rise with the siie 
of unit. For preliminary calculations a value of 0.94 may be assumed. Approximate el 



of various siie motors are given by Fig. 21. Efficiencies of pumps, fans, etc., will be found under 
theor raopective heading 

Combined Efficiency. The combined efficiency of a system of power generating, tran»< 
mitEiDn and tranaformation is the product (^ the various efficiencies of the apparatus used in 

the system. 



STEAM ENGINES 255 

One S'' oentrifugal pump, GKSwdiy, 1600 gal. per min.; total head, 100 ft; efficien<7, 0.65. 

One ventilating fan, capacity, 10,000 cu. ft. air i)er min.;^ 1" water, total preosure (5.2 lb. per iq. 
ft); efficiency, 0.40. 

One bucket elevator; capacity, 10 tons coal per hour; vertical lift, 40 ft; efficiency, 0.50. 

Five 10-hp. motors driving machinery. 

Efficiency of motors assumed as 0.80; generator efficiency, 0.94; electric transmission line Umb. 
5 per cent; machine efficiency of engine, 0.92; efficiency of sini^e belt drive, 0.94. 

« , . , , . 1600 X 8.33 X 100 

SohOunu dJip. of pump »» oon/vr^ w/tite — " ^21; electrical horsepower input to pump 

ooUUU X U.Od 

6.21 ^ 10000 X 5.2 

™°*^ ■■ /iTK* "■ S.26; dJip. of fan _ ., ^ oo/w^^ ■■ 3.94; electrical horsepower input to fan motor 

U./O' 0.4U X ooOUO 

3 94 ' 10 X 2000 X 40 

"■^S"*-^* «l~*ri«»l k<«eP<>wer input to elevator motor - o.60 X 0.75 X 60 X 33000 " ^-^^ 

5 XIO 
electrical h o rsep o wer input to motors driving machineiy » &■ 66.7; total electrical horsepower 

to be delivered to motors >■ 8.26 + 5.25 + 1-07 + 66.7 » 81.28; electrical horsepower at switoh- 

81.28 85.56 

board in power-house =■ = 85.56; Lhp. of engine »■ ^ «■ 98.93. 

Economical Temiliud Pressnre. TheoreticaUy, the steam should be expanded down to the 
back pressure carried to obtain maxiinum economy. As an engine is ordinarily designed to give 
maximum economy at normal load, thia is evidently impractical as the expansion line would 
necessarily fall below the back pressure line for the imderloads, resulting in an extremely poor 
economy for loads less than nonnaL 

Fortunately, in this respect, maximiim eoon<»ny, particularly with single-cylinder engines, 
is obtained ^len the terminal pressure is considerably higher than the back pressure. 

The results of tests made by O. H. Bonus ("En^e Tests," 1900) to determine the tenninal 
pressure th&t gives best economy for various classes of engines were as follows: 

TABLE 4 



^rpe <rf Engbie 



Sfanple riide-vslve cngfaw, non-eoodrnMing. 
Sfanpb iUde-vslve snginf, condendiiff . . . . 

Sfanpib Cortta engiiMS, mm-eondeniliig 

Sin^ CoriJM sngitif, eondenrinc 

Conpoand engiiMS, noD-ecmdeoiiiic 

CouqMoiid cngfaiM, eondenrinc 



Absolute Tenniiisl 

PreMure 

Lb. per Sq. In. 



80 to 40 
25 to 80 
20 to 26 
16 to 18 
18 to 22 
8to 6 



Determination of the Size of Simple Engine Cylinders. The power to be delivered by the 
engine having been previously determined is stated in brake horsepower if the machine is to 
drive by means of a beft or ropes. For direct-connected service the output at the generator 
terminals is stated in kilowatts (kw.) for direct cuirent and kik>-volt-amperes (k.vui.) for 
aHemating-current machines. 

The kilowatt output for alternating-current machines is the product dL the volts X 
amperes X the power factor. The power JadoT is the ratio of the apparent watts, as deter- 
mined from the readings of the voltmeter and ammeter, to the actual watts, and varies 
according to the class of load carried by the machine. 

The following approximate power factors, as given by the Qeneral Electric Co,, may be 
used in making preliminary estimates. 

* EffldHiey of motor and belt drive. 



1 



256 



POWER PLANTS AND REFRIGERATION 



0.95 when the load is principally made up of synchronous mot<»a and rotary converters. 

0.90-0.95 when the load is principally incandescent lamps. 

0.85 lighting and induction motors. 

0.80 induction motors only. 

For example, an A.C. machine rated at 1000 k.v.a. when used to supply current for inductioD 

motors only would have a rated output of 1000 X 0.80 or 800 kw. The efficiency (J^i) of a 

generator, as previously stated, is the ratio of the electrical hcMisepower output to the brake 

hcn^epower input. The machine efficiency of generators rated from 75 to 300 kw. may be 

assumed as 0.94 at ncmnal load in preliminary estimates. The machine efficiency (^i) for 

simple engines may be assumed as 0.92 at normal load and 0.90 for compound engines. As 

volts ^'C unDeres 

B kilowatts and 0.746 kw. is the equivalent of one electrical horsepower, the 

kw. 



1000 



indicated horsepower QJip.) required is equal to 

kw. 



or 



i.hp. 



0.94 X 0.92 X 0.746 



EiXEtX 0.746 
1.55 kw. 



The process followed in the determination of the cylinder dimensi<»8 will be made dear 
by the Examples (1) and (2) which follow. 

In practice, the process is usually a check on the dimensions as submitted by Uie engine 
builder covering the particular case at hand, as illustrated by Example (2) 

In case the engine exhaust is to be used for low-pressure heating a back pressure of approxi- 
mately 4 lb. should be used in the calculations. If the vacuum system of heating is to be em- 
ployed a back pressure of 2 lb. should be sufficient. 

TABLE 5 
XTSUAL PISTON SPEED AND REVOLUTIONS PER MINUTE FOR VARIOUS LENGTHS OF STROKE 



Stroke, Indm 


R. P. M.* 


Piston Speed, 
Feet per Mimxte 


8 


860-400 

800-880 

800-820 

280-800 

260-290 

250^280 

240-260 

220-240 

200-280 

126* 

120 

110 

100 

90 

80 


468-588 


9 


4ov o9o 


10 


600-688 


11 


618-660 


12 


680-680 


18 


641-606 


14 


660-608 


16 


688-640 


18 


600-690 


24 


600 


80 


600 


86 


660 


42 


700 


48 


720 


64 


720 







* Enginee equipoed with releaeinff ^rpe of geer are not <ntlinarfly operated at m Q>eed eiceedinf 125 r.pjn. 



TABLE 6 

SPEED OF ALTERNATING-CURRENT GENERATORS FOR DIRECT CONNECTION TO HI0H-8PKSD 

ENGINES 



K.VA. Rating. 
Speed r.p.ni . . 



tVoltage— 8-phaae. 
Voltag»— 2-phaae. 



60 
800 



76 
276 



240 
240 



106 
257 



480 
480 



160 
225 



600 
1150 



240 
200 



1160 
2800 



800 
160 



2800 



600 
120 



t Standard. The i peede vary aomewhat with different manufaeturers. 



STEAM ENGINES 



257 



Tbe direct coirent 125 volt exciters for the above machineB will have a oi^Muaty about as 
foDows: 

GoieratcNr Rating, k.v^ 60-75-105-150-240-300-600 

Exciter Rating, kw 4-5_ 7« g- 9-10-15 

TIm exciters are ordinarily belt driven from the engine and generator shaft, but may be 
obtained for direct connection to the engine shaft if desired. 

TABLE 7 
SPEED OP DIRECT-CURRENT GENERATORS FOR DIRECT CONNECTION TO HIGH-SPEED ENGINES 



Kw.R«tiiig*. 



20 


80 


40 


60 


60 


76 


100 


126 


160 


200 


260 


860 


860 


800 


276 


276 


276 


200 


260 


260 


200 


200 



800 
160 



^Standard vottasa 110 and 220. The speeds vary aomewhat with different maaufaeturera. 

Steam Pressures Employed. Plants in which simple engines are used for the prime movers 
are now ordinarily designed for a steam pressure of 100 to 125 lb. gage. With compound engines 
a I^esBure of 125 to 150 lb. gage is usual, while in steam turbine plajits a pressure of 150 to 200 
lb. gage is customary. 

Cut-off at Normal Ratiiig. It is customary practice to base the normal rating of simple 
noorocmdaising engines on a cut-off of ^ stroke or 4 expansions for the usual initial pressures 
empbyed. For single-valve engines this gives about the most economical terminal pressure 
as stated in Table 4. 

Biample. Initial pressure 125 lb. gage or 140 lb. absolute terminal pressure with four expansions 
18 140/4 or 35 lb. 

Fw compound non-condensing engines a terminal pressure of 20 lb. may be used and 10 lb. 
for compound condensing engines. Based on the above figures Tables 8 and 9 have been 
calculated: 

TABLE 8 

MEAN EFFECTIVE PRESSURES FOR SIMPLE ENGINES 

Diagram Faetor 0.80. BaekPreinare 16.7 Pound 



Initial 

Above Throttle 



QtCB 



86.8 
90.8 
96.8 
100.8 
106.8 
U0.8 
116.8 
120.8 
126.8 
130.8 
186.8 



Abaolttte 
ft 



100 
106 
110 
116 
120 
125 
180 
186 
140 
145 
160 



KCut-off 
r -4 



25.0 
26.2 
27.5 
28.8 
80.0 
81.8 
82.6 
88.8 
86.0 
86.8 
87.6 



42.9 
44.9 
48.9 
61.9 
64.9 
67.9 
60.8 
68.8 
66.8 
69.8 
72.6 



84.8 
86.0 
89.1 
41.6 
44.0 
46.8 
48.6 
61.0 
68.4 
66.8 
68.1 



HCutHiff 
r »8.88 



80.8 
81.6 
88.0 
84.6 
86.0 
87.6 
39.0 
40.6 
42.0 
48.6 
46.8 



49.4 
62.7 
66.0 
69.8 
62.6 
66.9 
69.8 
72.6 
76.8 
79.2 
82.6 



89.6 
42.2 
44.8 
47.4 
60.1 
62.7 
66.4 
68.0 
60.6 
68.4 
66.0 



26% Overload 



60% Overload 



CutK)fl at Normal Load 



H 



42.9 
46.0 
48.9 
61.9 
66.0 
67.6 
60.8 
61.8 
64.2 
69.7 
72.6 



H 



49.8 
62.7 
66.0 
69.2 
62.6 
66.9 
69.2 
72.6 
76.8 
79.2 
82.6 



I 



3 



H 



61.4 
64.0 
58.6 
62.2 
66.0 
69.4 
72.9 
76.6 
80.1 
88.7 
87.2 



I 

fa 
& ■ 

V 



H 



69.8 
68.8 
67.2 
71.1 
76.2 
79.1 
88.1 
87.0 
91.0 
96.1 
99.0 



Overload Capacities. For direct-connected units both engines and generators, as ordinarily 
rated and designed, are capable of handling continuous overloads of 25% and 50% momentarily 
without undue heating of parts. 



258 POWER PLANTS AND REFMGERATION . 

Simple automatic cut-off engines as ordinarily rated are designed for a cut off of about ){ 
stroke at normal load, and for a 50% overload the cut oS must be increased to about Vio of 
the stroke, the maximum cut off obtainable being approximately "^/lo. 

The slow and medium speed Corliss type of engine must be equipped with a double eccen- 
tric gear to carry a 50% overload, as the limit of cut off with a single eccentric gear is approzi- 
matdy }i stroke. The maximum overload capacity of the prime mover and the regulation 
required should be clearly stated in the specification and proposal. 

Bxample. (1) Determine the cylinder dimensions for a non-oondensing simple engine for direet 
connection to a 50 kw. direct current generator (normal rating) for the following conditions of operatloo: 

Steam pressure at engine throttle valve 100 lb. gage. Assumed back pressure 2 lb. gage. 

Speed from Table 7 is 275 r.p.m. cut off K stroke for normal load. 

The theoretical mean forward pressure ^ 0.5965 (100 + 14.7) « 68.4 lb. per sq. hi. 

The diagram factor. Table 2, is 0.80. Length of stroke from Table 5 is 12^'. The Lhp. «* 50 X 

1.55 s 77.5. The expected mean effective pressure (m.e.p.) » 0.80 (68.4 » 16.7) » 41.4 lb. per 

sq. in. 

.- _ . i.hp.X 33000 77.5X33000 ,,_ , a- ji- 

Area of the piston A = ^ , ^, = » 112 sq. in. ooireeponding di- 

P^LN 41.4 X 1 X (2 X 275) »-«— • 

ameter = 12'^ The engine would be made 12" x 12'^ 

Bxample. (2) The size of cylinder for a 100 kw. unit is given by Table 14 as 14'' x 14" at 250 
r.p.m. initial pressure 120 lb. gage or 135 lb. absolute. The required i.hp. » 1.55 X 100 or 155. 

The theoretical mean forward pressure by calculation is 94 lb. for )^ out off or 3.33 expansions. 

The expected m.e.p. is 94 X 0.80 (diag. factor) - 16.7 » 58.5. A » area 14" dia. C3^inder or 153.9 
sq. in. L «s length of stroke or 1.166 ft. 

rru 1 1 *^ 1. • ♦!, r 58.5 X 1.166 X 153.9 X (2 X 250) ,^ 

The calculated i.hp. is therefore or 159. 

33000 

This engine, at normal load, will require approximately 155 X 28}^ or 4420 lb. of steam per hoar. 
See economy curves, Fig. 25. 

Overload Capacity, If this engine is to carry a 50% overioady the m.e.p. at normal load must be 
increased by this amount. 

The m.e.p. required at maximum load is therefore: 1.5 X 58.5 ■■ 87.7 lb. per sq. in. 

/ 1 -f log, r \ 1 -f log, r 
Then 0.80 ( 135 X - 16.7 J « 87.7. - 0.94. The corresponding vahie erf 

r from Table 1 is 1.50, the cut off being 0.66 of the stroke, which is within the limit of cut off for 
single- valve automatic engines. 

CX)MPOUND ENGINES 

The loss incident to cylinder condensation is. due to the condensing out of a smaU portioQ 
of the steam when brought into contact with the colder cylinder wall. The cylinder wall is 
maintained at a fairly constant temperature, approximately a mean between the initial and final 
temperature of the expanding steam. 

It is evident that the greater the ratio of expansion the greater will be the temperature 
range of the steam in the cylinder, the lower the temperature of the cylinder wall and ocmsequently 
the greater the loss will be from this cause. If the temperature range is reduced by dividing 
the expansion between two cylinders it is found that the loss is much less than when a sin^^e 
cylinder is used, giving a gain in economy of approximately 10% for non-oondensing units, as 
will be noted by a ccnnparison of the water rate curves (Fig. 25). Were it not fcnr the reduc- 
tion in the heat loss there would be no reaacm tot compounding which would be sufficient to 
warrant the expense. 



STEAM ENGINES 269 

In the eompound engine two cylinders are provided. The steam is first admitted to the 
ki(^'pre9sure cylinder and expanded to several times its (»iginal volume, the terminal or final 
luneBBure in Uiis cylinder being approximately one-half to one-third of the initial pressure. This 
expanded volume, at the reduced pressure, is then passed to a second or low-pressure cylinder 
and the expansion completed. 

ThecM^cally, the total expansion is the same as if carried out in the low-pressure cylinder 
with the same volume of steam and at the same initial pressure as admitted to ^e high-pressure 
ey^nder. 

Coaibln«d Indicator Diagram* The action of a ccnnpound engine is conveniently studied 
by means of the theoretical combined indicator card or diagram abcfghk (Fig. 22). 
An fwesBures are absolute lb. per sq. in. and all volumes are stated in cubic feet. 
Let pi » initial absolute pressure. high-i>ressure cylinder. 
Ps ™ terminal pressure high-pressure cylinder. 
Pr " back pressure high-pressure cylinder, 
e receiver pressure. 

B initial pressure low-pressure cylinder. 
ps » tenninal pressure low-i>reesure cyliixier. 
Pt a back pressure low-pressure cylinder. 
a = aiea high-in:eeBure cylinder, sq. in* 
A » area low-pressure cylinder, sq. in. 

a 

J- » E = cylinder ratio or ratio of cylinder volumes and also areas, when the strdces 

are made the same length as is cusUHnary practioe. 
V\ » volume high-i>re8sure cylinder. 
Fs » volumclow-pressure cylinder. 

Hien — « — — B 
Vt A 

T "3 total ratio of expansion » — 

n "" ratio of expansion in high-pressure cylinder » — 

V* 

Ti = ratio of expansion in low-pressure cylinder = -^ 

Pi 

Cut €M in high-pressure cylinder = . 

R?\r 

m.e.p. h.p. cyl. (diagram abck) = pi ( ^ ] ^ Pr 

m.e.p. Lp. cyl. (diagram h^gh) ^ Pr \ ) ~ P« 

m.e.p. of h.p.X a + m.e.p. of Lp. X A ^ work per foot of stroke both cylinders. 
But a - AR, 

Then ii (m.e.p. of h.p. X i2 + m.e.p. of l.p.) » work per foot of stroke, both cylinders 
Fot equal work developed in each cylinder m.e.p. of h.p. X R ^ m.e.p. of Lp. 

When the cut oS. for the low-pressure cylinder is such that the volume displaced by the low- 
prenure piston at the point of cut off is equal to the volume of the hi^ pressure cylinder, then 
|>s a TV. If the cut off in the low-pressure cylinder is lengthened so that the volume drawn 
from the receiver is greater than the high-pressure cylinder voliune there occurs a drop, free 
expansion, in pressure from p^^ to jv at release in the high-pressure cylinder, as shown by cards 
(Acti and idfgh. The effect of lengthening the cut off in the low-pressure cylinder results in an 



260 



POWER PLANTS AND REFRIGERATION 



increase in the m.e.p. of the high-pressure cylinder, with a corresponding decrease in the iii.e.p. 
of the low-pressure cylinder. 

The examples following serve to illustrate the principles involved. 

Calculations for the Cylinder Dimensions for Compound Engines. The cylinder ratio or 
the ratio of the volume of the hi^-preesure to the volume of the low-pressure cylinder is usuaDy 
made such that the work is approximately equally divided between the two cylinden at Donnal 



^JCtttotrH.P.Qyl, 



H.P. 




I 



Pr-^8 



^^ 



■Vr 



Vol, H.P.Qyl* 



^ /? — 26 



:C. 



Pr. UnB 



Pa- 



4^ 



YoK LP.Qyf, 
FlO. 22. OOMBINBD PBaBSUBB-yOLUMB DiAOBAM 1X>B A OCMPOUHtD BxnHMB. 



'Kll 



or rated capacity. For non-condensing engines this ratio is usually made 1 to 2^ or 3 and 
for condensing machines 1 to 3 H or 4. The division of the load between the cylinders may be 
made by adjusting the length of cut-off in the low-pressure cylinder. As the low-pressure elit- 
es is lengthened the lower becomes the receiver or back pressure on the high-pressure eylmder, 
increasing the work done by this piston and decreasing the woric performed in the low-pressure 
cylinder. 

In cross-compound engines of standard construction the strokes of the high and low-prea- 
sure cylinders are made the same and are of necessity the same tat the tandem type. Therefore, 
having decided upon the stroke and area of the low-pressure piston the area of the high-pressure 
piston is obtained by dividing the area of the low-pressure piston by the cylinder ratio. 

The area of the low-pressure piston is calculated on the assumption that aU of the work la 
to be performed in this cylinder in the same manner as for the case of the simple engine. That 
this assumption is correct is evident from an inspection oi the combined indicator card (Fig. 22). 
If the intermediate line (receiver pressure line) is removed the card would be that for a simple 
cylinder of the same volume of the low-pressure cylinder. 

Hence if the same initial pressure is used in a cylinder of the sise (A the low-i>re8Bure cylinder 
and if the same total ratio of expansion be used, this cylinder, theoretically, will develop the same 
power as the compound engine. 

Usual A89umption8. It is usual to assume about the following absolute terminal and bade 
pressures in compound engine calculations for the normal or rated load. Terminal pressure pi 
a 20 to 25 lb. for non-condensing and 10 lb. for condensing engines. 



STEAM ENGINES 



261 



Back presBure p» » 16.7 lb. for nwwxwideniring and 2 to 3 lb. for oondenBing engmee o(mv 
to a 25.02 in. and 23.88 in. vacuum. 



If the engine is to be able to oarry a 50 per cent overload the cut off in the high-preesure 
ejrlinder should be estimated in advance f (nr the overload and kept within the limit of the range 
of cut off of the vahre gear. The foUowing example illustrates this ^int: 

TABLE 9 

MEAN KKFEC T IVE PRESSURES FOR COBfPOUND ENGINES 

NON-CONDBNBINO 

PI -> 20, Back PreHon pi - 16.7, Diagram FMtor - 0.76, Cylinder Ratio 12 -> 1 : 2.6 



INRUL I 


^mmmmm 












60% OVBBLOAD 






Ratio 
of 


Cutoff 
H.P. 


Terminal 
Prearare 


Theoret- 
ical 


Eixpeeted 

mIe.p. 












Gafa 


Abaolnta 

pi 


f 


CyUiMkr 


P9 


M.E.P. 




Eixpeeted 
aI>£!*P« 


Cutoff 
H. P. CyL* 


190.8 


186 


6.76 


0.870 


20 


41.6 


81.1 


46.6 


0.600 


126.8 


140 


7.00 


0.867 


20 


42.2 


81.6 


47.4 


0.690 


ISO.8 


146 


7.26 


0.844 


20 


42.9 


82.1 


48.2 




186.8 


160 


7.60 


0.888 


20 


48.9 


82.7 


49.0 




140.8 


166 


7.76 


0.828 


20 


44.2 


88.0 


49.6 




146.8 


160 


8.00 


0.818 


20 


44.9 


88.7 


60.6 


0.600 


160.8 


166 


8.26 


0.808 


20 


46.6 


84.1 


61.6 




166.8 


170 


8.60 


0.294 


20 


46.0 


84.6 


61.8 


• • • • • 


160.8 


176 


8.76 


0.286 


20 


46.6 


86.0 


62.6 


0.466 



91 » 10, Back Pr e ife 91 « 8, Diagram Factor « 0.76, Cylinder Ratio l{ - 1 : 4 



126.8 


140 




0.286 


10 


88.4 


26.0 


87.6 


0.66 


186.8 


160 




0.266 


10 


84.1 . 


26^ 
26.0^ 


88.8 




146.8 


160 




0.260 


10 


84.7 


89.0 


• • • • • 


160.8 


166 


16.6 


0.242 


10 


86.0 


26.2* 


89.8 


0.42 


166.8 


170 




0.286 


10 


86.8 


26.6 


89.8 




166.8 


180 


18. 


0.222 


10 


86.9 


27.0 


40.6 


• • • • • 


176.8 


200 


20. 


0.200 


10 


86.9 


27.6 


41.4 


0.88 



•Cat off In H. P. cylinder- s^. 

A X r 



»»awipU- Determine the cylinder dimensions for a oompoimd non-condensing engine direct 
conneeted to a 200 kw. D. C. machine the rated speed of which is 200 r.p.m. Initial pressure 125 lb. 
gage, atmospheric exhaust. 2 lb. gage back pressure. 

Assume a terminal pressure ps -• 25 lb., a diagram factor of 0.75, and a cylinder ratio R of 
1: 3.6. 



Tliis gives for the total expansion ratio r » 



140 
25 



5.5. The value of ^ "^ ^^^ '' from Table 1 for 



i 



r -> 60^ is 0.492. 

Expected m.e.p. r e f erred to the low-pressure cylinder » 0.75 (140 X 0.49 — 16.7) ■■ 39 lb. sq. in. 

F^om Table 5 an 18^' stroke will be assumed for the rotative speed to be employed L « 1.5 ft. 
^' -400. 

IJip. required is 200 X 1.55 or 310. 

A 1 ^ .1 310 X 33000 .-_ 

Area low-press, cylinder — ■■ 437. sq. m. 

Conesponding diameter 23 A''* 

For a cylinder ratio of 1 :2.5 the area of high-pressure cylinder is 4374-2.5 or 175 sq. in. Ck>rreBpond- 

£__ ii:-„i, ■# ■■- 1 Kff 
tng cnameter 10 . 

The engine may therefore be made 15'' X 24" X 18''. Ck>mpare with the cylinder dimensions 

given by Table 17 for this siae unit. This engine will not carry a 50% overload as shown by the 

iple following. 



262 POWER PLANTS AND REFMGERATION 

Cut Off in HiohrPreaaure Cylinder for Fifty Per Cent (herloaS Capaciiy. In order that the 
engine may cany an overload of 60% the m.e.p. at normal load must be increased by this amount. The 
m.e.p. required at mayimiim load for the engine in the preceding example is: 1.5 X 39 » 58.5 lb. 
per sq. in. 

r /I + log, r\ "1 ^ 

Then 0.75 I 140( j-16.7 I «58.5 in which 0.75 b the assumed diagram factor at 60% 



overload. 

1 + lo&.r 



— 0.676. The corresponding value of r from Table 1 is 3.3. 



The cut off required in the high-pressure cylinder is therefore — - X 2.5 (esd. ratio) or 0.76 of 

the stroke. 

This is bejrond the limit of cut off obtainable with automatic single-valve engines, and it would, 
therefore, be necessary to choose a lower terminal pressure for the normal load, thus obtaining a greater 
total ratio of expansion. If pi « 20 lb. abs., then r » 7 and the expected m.e.p. referred to the low- 
pressure cylinder is 31.6 (see Table 9). Area of the low-pressure cylinder is 540 sq. in. and the cor- 
responding diameter is 26K"> The cut off in the high-pressure cylinder is found to be 0.57, which is 
within the limit of the range of cut off for automatic single-valve engines. Area of high-pressure 
cylinder will be 540/2.5 or 216 sq. in., corresponding to a 16>^ in. diameter. 

Division of Woric between Cylinders. Assuming no drop in pressure exists at the end of 
stroke in the h.p. cylinder that is p, — jv we have the relation piVc » PxVi '^ PrV ^ ptVt 

Pi - 140 (nearly) p, = 25. i2 - ^* - -i" 

Vi 
Pi - P* — or p, - 25 X 2.5 - 62.6 lb. sq. in. 

mu 140 ^# ^ 62.5 ^^ 

Then n - ttz - 2.2 and u « -— r " 2.5. 
62.5 25 



Theor. m.e.p. high-press. cyL ™ 140 ( «/^ ) " ^^.5 « 51.3. 
Theor. m.e.p. low-press. cyL - 62.5 f ^ ' J — 16.7 = 31.2. 



For equal work in each cylinder m.e.p. of h.p. cyl. X R should equal m.e.p. of low-preasure 
cylinder. 

51.3 X — B 20.5 lb. per sq. in. It is seen that the low-pressure cylinder wiU be doing 

more work in this case so that a reduction in the receiver pressure is necessary to obtain 
an equal division between the two cylinders. Try pr * 48^ then the 

Theor. m.e.p. of h.p. cyl. = 140 ( ^^ j — 48 = 65.8, ^« = ^ " l-^- 

/I -f log, 1.9\ 
Theor. m.e.p. of l.p. cyL = 48 ( ~ 1 - 16.7 = 24.5. 

Then 65.8 X t— (R) = 26. m.e.p. of high-pressure cylinder referred to the low pressure. 
2.5 

Approximately equal work will therefore be performed in each cylinder for the assumed con- 
ditions of operation. 

The reduction in receiver pressure is obtained by lengthening the low-pressure cut-off as 
shown. 



STEAM ENGINES 



263 



Range in tiie Cylmden. 

Temperature oorreqpondiiig to 140. lb. is 353.1^ F. 
Temperature oorreQwnding to 48. lb. is 278.4^ F. 
Temperature oorreQwnding to 16.7 lb. is 218.*" F. 
Temperature range h.p. cylinder is 353.1 — 278.4 • 
Temperature range Lp. cylinder is 278.4 — 218. < 



74.7* F. 
60.4** F. 



STANDARD OF PERFORMANCE FOR STEAM ENGINES AND TURBINES 

The HnnMne Cycle. The Rankine or Clausiua cycle is quite generally accepted as a stand- 
ard of oompariaon for the performance of steam engines and steam turbines. In this ideal cycle 



T 
I 
I 
I 
I 



I 
I 

r 
I 



[3 ? 




Vo/ume^ 



Fio. 23. Rankinb's Ctclb. 



steam is assumed to be eiqumded adiabatically in a non-conducting cylinder without clearance, 
the escpansion bdng caiiied down to the back-pressure line. The diagram (Fig. 23) represents 
the aotioQ of steam or other vapor when operating on the Rankine cycle with complete ex- 
pansUni. 



Pi, Pi 

Fi, F, 

9i 
9t 

Pi 
Pt 

ri,r% 
A 



initial and final absolute pressure, lb. per sq. ft. 
q)ecific volume of saturated steam corresponding to Pi and P%. 
q)ecific volume of the Uquid corresponding to Pi and Ps. 
initial and final volumes, cu. ft. 
heat of liquid corresponding to initial pressure. Pi. 
heat of Uquid ctHresponding to final pressure, Pj. 
quality of vapor admitted to cylinder, 
quality of vapor exhausted, 
internal latent heat, initial state, 
internal latent heat, final state. 

increase in the volume per lb. in changing frcmi a liquid to the vapor state corre- 
sponding to pressures Pi and Pj. wi « Vi" — th', wi — »i" — »/. 
total heat of vaporization corresponding to pressures Pi and Ps. 
1/778 reciprocal of the mechanical equivalent of heat. 
Pi + APiUi. r, « />, -h APtUi, 



If a vapor be used in a non-condensing cylinder without clearance and the expansion car- 
ried down to the back pressure line we would obtain a diagram giving the relation between the 
pressure and volume similar to the ideal diagram (Fig. 23). 

The vapor is admitted to the cylinder, at constant absolute pressure. Pi, from a to 6. At 
b cut off occurs and the vapor expands adiabatically to c, at which point the exhaust opens and 
the vi^xMr is exhausted at a constant absolute pressure P%. 



264 POWER PLANTS AND BEFRIGERATION 

For one pound of yapw admitted to the cylinder the work performed by the vapor oo the 
piston from a to 6 is: 

PiFi - PiMt^" -O +t^1 = Pi(^it*i + t^') (1) 

For expansion from 6 to c the heat that is changed into work is 

= 778 (qi - gi + XiPi - xtpt) ft.-lb (2) 

The work performed by the v&por on the piston from c to d is 

PtVt « P, [x, (r," -V) +vt'] «= Ptixiut + vt') (3) 

The net work W performed on the piston by the vapor is (1) + (2) — (3) or 

W - 778 Iqi + XiPi + APiXiUi - gj - xtpt - AP^ittUt + A(PiVi' - P^,')l . (4) 

Since the last term is small it may be dropped without appreciable error. Substituting 
ar = X (p + APu) in the above equation, we have W — 778 (gi + Xin — gi — xiTi) ft.-lb. per 
pound of vapor suppUed. 

Let Qi " total heat per lb. initial state. 
*" gi + Xiri saturated vapor. 
Qt » total heat per lb. final state. 

= gi + xifi. 

Then W «= 778 (Qi — Qt) ft.-lb. of work performed per lb. of vapor supplied. 

If the vapor is initially superheated then Qi « /f i + Cpt(t — t,) in which Hi «* the total 
heat of saturated vapor per lb. = gi -h n, 

Cpt B mean specific heat of superheated steam. (See "Mean Specific Heat Curves" in the 
Chapter on "Water, Steam and Air.") 

t s temperature of the superheated steam. 
ts = temperature of saturated steam corresponding to pressure Pi. 

The determination of the final quality Xt, citer adiabatic expansion has taken place, in- 
volves the use of the quantity entropy. Entropy, as defined in the Chapter on " Water, Steam 
and Air, " is the ratio of the heat added to a substance divided by the absolute temperature 
at which it is added, values for which are given by the steam tables. 

When expansion or contraction of a gas or vapor takes place without loss or gain of heat 
the change is said to be adiabatic, therefore the entropy remains constant for such a change. 

This affords a means for determining the final quality Xt and therefore Qt as follows: 

If the vapor is initially dry or wet saturated, the necessary data will be found in the satu- 
ated steam table. 

Si, St » entropy of the liquid corresponding to the initial and final states. 

Xjri XfTt 

<=-, -^ = entropy of the vapor, initial and final states in which Ti, Tt * the initial and 

Ti Tt 

final absolute temperatiu'es. 

XiTi Xfrt 

Then Si + -^ — St + -^ from which the value of xt may be determined. 
Ti Tt 

If the v£4x>r is initially superheated the entropy of the superheated vapor may be rou^y 
approximated by the following method. 

The heat added due to the superheating « Cps{t — t,) and the averoife absolute tempera- 

* u'u,' AA^' rp (< -f 460) + {Is + 460) 

ture at which it is added is Ta = . 

2 

The entropy of the superheated vapor may therefore be stated as. 

5.-5.+—+ ^—. 

The above approximate method involves an error which, however, is not particularly serious 
when the superheating does not exceed 200 degrees. 



STEAM ENGINES 265 

For exact values oi Sisee OoodenougVa tables for superheated steam. 

If the vapor is initially wet or saturated we have the relation Si =« St + "zr ^^om which 

the value oi xt may be determined. 

All problems involving adiabatic expansion are solved with rapidity and sufficient accu- 
racy by means of the MctUer chart or diagram in the Chapter on ''Steam Turbines." 

The chart has largely supplanted the steam tables in this connection. 

Bzample. Required the amount of heat changed into work per pound of steam supplied an engine 
or turbine wcn'king on the Rankine cyde. Initial pressure »> 125 lb. absolute, lOO** superheat with 
an exhaust or terminal pressure of 2 lb. absolute. 

t, - 344.4° F. and « - <, + 100 or 444.4® F. 

The average specific heat of superheated steam for the range of temperature stated is 0.556. 

The heat added for superheating is 100 X 0.556 » 55.6 B.t.u. and the average absolute temper- 

344 + 444 

ature at which it is added is r + 460 - 854®. The increase in entropy due to superheating 

2 

55.6 Tx 

Is — — or 0.065. The entropy of the liquid is 5' - 0.4950 and for vaporisation is — - 1.0908. The 
854 i 1 

entiopy of superheated steam under the conditions stated is iSi * 0.495 + 1.091 + 0.065 -• 1.651 or 
from Ooodetujiugh*9 tables direct is 1.651. 

Xt Tt 

The entropy for the final condition is St — jSs' + ~~;;r~'* As the expansion is assumed to take place 
adiabatically Si ■■ jSs. 

For 2 lb. pressure ^ - 1-7452 8t' - 0.1750. 

Then 1.65 » xt X 1.7452 + 0.1750. xt - 0.85 + 

The heat that is changed into indicated work per pound of steam used is the difference between 
the heat received by the engine Qi ** 9i + ri + Cp {t — U) and that rejected, or Qt -■ X}grt + 9s* 

Therefore TT - 778 [^i + ri + C^ « - <*) - x«ri - qt] or TT - 778 (315.1 + 876.9 + 0.556 
X 100 — 0.85 X 1022.2 — 94.0) » 221,520 ft.-lb. per pound of steam supplied. From this must be 
subtracted an amount equivalent to the losses that occur in the actual engine to obtain the work 
delivered by the crank shaft. 

EFFICIENCY STANDARDS 

There are two standards used in estimating or making comparisons of engine and steam 
turbine performance. 

Thennal Effidancy. The transformation of heat energy into mechanical work is always 
accompanied by an unavoidable loss. 

For example, heat is supplied a boiler and as a result water is evaporated into steam. The 
heat that it is necessary to supply to produce one pound of steam is the sum of heat required 
to raise the temperature of the feed water from its initial temperature to the temperature corre- 
sponding to the pressure existing in the boiler plus the latent heat of vaporization corresponding 
to this pressure. 

Tlie steam generated is used in an engine or turbine and the same weight of steam at a lower 
pressure and heat content is exhausted or rejected by the machine. The latent heat of the 
exhaust steam, in so far as the engine is concerned, is a direct loss inasmuch as thd^ is no means 
of tomsforming it into useful work. A portion of the latent heat, however, may be utilized in 
raising the temperature of the feed water, supplied the boiler, up to practically the temperature 
of exhaust steam, and is so considered in this connection. It is evident that the less the weight 
of steam required to perform a definite amount of work the more efficient the engine is as an energy 
oonversioii medium. 



266 POWER PLANTS AND REFRIGERATION 

The steam consumption or water rate of an engine or tm'bine is also used as a means of com- 
paring the economic performance of this class of prime mover. The thermal efficiency of an engine 
is defined as the ratio ol the heat converted into work in the steam cylinder per pound of steun 
supplied to the heat supplied per pound of the steam. 

Let Qi = heat of the liquid corresponding to the initial pressure of the steam at the 
engine. 
qt = heat of the liquid corresponding to the pressure and temperature of the exhaust 

steam, 
n — llitent heat corresponding to the initial pressure. 
Xi = quality of the steam supplied engine. 
Then Xin + qi — qt = heat supplied, per pound, to the steam as used by the engine. 

W = weight of water used per indicated horsepower developed per hour (i.hp.-hour). 
1 B.t.u. = 778 ft.-lb. 1 hp. « 33,000 ft.-lb. per minute. Therefore 1 hp.-hour is the 
equivalent of ^• 

33,000 X 60 /rrrr^ , 

^^ ^, or(2646 B.t.u. per hour. 

778 v\ ^-- ^ 

The heat transformed into useful or Indicated work in the steam cylinder per pound of 
steam or water supplied is equal to 

^ 2546^^ 
Oa = -^ B.t.u. 

The thermal efficiency of the actual engme according to the definiti<m given is th^nefOTe 

2546 A ■ ■■' 



a 



WixiTi + gi - qt) 



Example. The steam consumptioD of a certain engine is {W) 35 poimds per i.hp.-hour when 
supplied with dry and saturated steam at 120 lb. per square inch gage with atmosphere ezhau8til4.7 
lb. per sq. in. absolute) xi » 1, fi + 9i -* 1193.2; 9s » 180. The thermal efficiency of the actual engine 
is therefore: 

2546 
^- - 35 (1193.2 - 180) - »•""» ">' ^•2%- 

The thermal efficiency of the ideal Rankine engine is similarly defined as the ratio of the heat 
converted into work in the steam cylinder per pound of steam sufH>lied to the heat supplied per 
pound of steam. In the Rankine engine the expansion takes place adiabatically. 
Let Qi = heat content of the steam supplied per pound. 
= xin -f qu 
Qi — heat content of the steam rejected per pound. 
= XiTt -h qt' 
Knowing the initial and back, or exhaust pressures and the numerical values of Xiri and ^, 
the theoretical final heat content Qt is readily determined by means of the MdUier chart or the 
entropy tables. 

Qi — <?i = heat converted into work per poimd of steam supplied in the ideal Rankine Angip ^*. 
The heat supplied the engine per pound of steam is as before given or Xin + gi — gi. 
The thermal efficiency of the ideal Rankine engine is therefore: 

Ntf = 



Xiri -f gi - gi 



The steam consumption per i.hp.-hoiur or water rate of the ideal Rankine engine is: 



r 



r, 



STEAM ENGINES 



267 




■^ 



-^ 



80 90 100 110 
Indicated Horse Power 




W^Corfcfd Water RaU 
Qt-Heat Content of Steam Supplied per lb. =1093 
igf=Heat Content of Steam at Completion of 
Adlabatlo Expanafon^lOSO 



30 40 50 60 



70 80 90 100 110 

Indicated Horse Power 

Fio. 24. 



120 ISO 140 150 



268 



POWER PLANTS AND REFRIGERATION 



Bxample. The thennal effioienoy of the ideal Rankine engine working between the same tirtiawiio 
limits as given in the preceding example, may be calculated as follows: 
Qi » 1193.2. From the MoUier chart Qt - 1035 (approximately). 
1193.2 " 1035 



Then Nr - 



- 0.156 or 15.6%. 



1193.2 " 180 

The theoretical steam consumption of the ideal Rankine engine for the given oonditions 

2546 2546 



TF« - 



Qi - Qt 1193.2 - 1035 



« 16.1 pounds per i.hp.-hour. 



TABLE 10 

STEAM CONSUMPTION, IDEAL RANKINE CYCLE 
Pounds of Steam p«r Hp^Hour 





Pressure 

Pi 

Pounds per 

Square Inefa, 

Absolute 


Saturated 


SUPBBBB4T 




Steam 


60* P. 


100* F. 


200«F. 




Bsek Pressure pt. Pounds per Square Inch. Absolute 

• 




16 


2 


16 


1 


16 


1 


16 


1 


80 


20.8 
19.6 
18.6 
17.6 
16.8 
16.2 
16.7 
16.2 
14.8 
14.4 
14.1 
18.8 
13.6 
18.1 
12.6 


10.6 
10.2 
9.9 
9.7 
9.6 
9.8 
9.1 
9.0 
8.9 
8.8 
8.7 
8.6 
8.6 
8.8 
8.1 


20.2 
18.9 
17.8 
17.0 
16.2 
16.6 
16.1 
14.7 
14.8 
18.9 
18.6 
18.8 
18.0 
12.6 
12.0 


8.9 

8.7 

8.6 

8.8 ' 

8.2 

8.0 

7.9 

7.8 

7.7 

7.6 

7.6 

7.4 

7.8 

7.2 

7.1 


19.8 
18.1 
17.1 
16.8 
16.6 
16.0 
14.6 
14.1 
18.7 
18.4 
18.1 
12.8 
12.6 
12.1 
11.6 


8.8 
8.6 
8.8 
8.1 
7.9 
7.8 
7.7 
7.6 
7.6 
7.4 
7.8 
7.2 
7.1 
7.0 
6.9 


17.6 
16.6 
16.7 
16.0 
14.4 
18.9 
18.6 
18.0 
12.7 
12.4 
12.1 
11.9 
11.6 
11.2 
10.8 


8.1 


90 


7.9 


100 


7.8 


110 


7.6 


120 


7.4 


180 


7.8 


140 


7.2 


160 


7.1 


160 


7.0 


170 


6.9 


180 


6.9 


190 


6.8 


200 


6.7 


220 


6.6 


260 


6.6 







Potential Efficiency. The thermal efficiency, however, is not generaOy considered as satis- 
factory a basis for making comparisons of performance of steam motors as another ratio tenned 
the potential efficiency or efficiency ratio. 

Indicated potential efficiency is defined as the ratio of the heat ccmverted into indicated 
work in the actual engine to the heat converted into indicated work by the ideal Rankine engine 
working between the same pressure limits. We have here a ratio which expresses the degree 
which the actual engine transforms into woric the heat that might possibly be converted into 
work by a perfect engine. 

2546 

Potential efficiency AT^ » -2- = or —-— -— 

Qr Qi-Qt W(Qi-Qt) 

The most useful criterion and one which takes into account the engine friction as well is 
known as the '' brake potential efficiency " or 

2546 

* Wt(Qi - Qt) 

Wb being the steam consumption in pounds per brake horsepower of the actual engine. As we 
have no means of determining the indicated h<n«epower of a turbine the " brake potential effi- 
ciency " is one that is frequently made use of in this connection. 




7 ^ 

^^^^ , 'Sfpam Turbines 



'i 



Copocif/. 
K.W- 


JfiaomGsnsampllaf: IJfxper M.IVMxJT 


LO€7Cf 


i 


/ 


/4 


3S 


6S 


63 


50 


7S 


72,5 


5^.5 


575 


/OO 


60.5 


4^.5 


45.5 


/ZS 


66 


56 


56 


300 


57 


53 


5?.5 



Ah^a:- /hOuc^ a/prax. 

//7 kfi)/e fisreirery J(?* 
/ncrease in Jhe /h//n7/ 
pressure. 

/Icfcf ZJ(> fhr each /¥> 
ofnro/s/ure /r? ^s/eaim. 
/ncreox tw^rofebfrifm^- 

ores skf/ecf^ Z^3%:, ^-5%} 
4^''7i%; 5^-/oi^, 



Fio. 25. 



270 



POWER PLANTS AND REFRIGERATION 



Example. The potential efficiency of the actual engviCi the data for which are given by the preced- 
ing examples, ia: 

2546 ^ 

^'-36(1193.2 -1036) -^•^^^^^-- 

That is to say this particular engine cylinder transforms into useful indicated work 46% of the amount 
that is theoretically possible to transform in a perfect engine working on the Rankine cycle for complete 
expansion between the same pressure limits. 

Steam Consumptioii or Water Rate. The most generally used measure of tlie eoonoznic 
performanoe oi steam engines and turbines is the weight of dry steam, as determined by tests, 
used per hour per unit of power developed by the machine. This is frequently termed the tcaier 
rate of the machine. The water rate is ordinarily determined by weighing the feed water for 
non-condensing engines and the condensate for condensing engines. 

For recq3rocating engines the steam consumption or water rate is ordinarily stated as the 
weight of dry saturated steam supplied the engine per indicated horsepower p^ hour. 

For steam turbines the steam consumption is stated as the weight of steam supplied per 
brake horsepower-hour or per electrical horsepower-hour or kilowatt-hour. The power devel- 
oped per unit weight of steam supplied varies with the initial pressure, initial quality or super- 
heat and the back pressure or degree of vacuum maintained. It is therefore essential to state 
these conditions in every case, in order that a comparison may be made and that the data may 
be used properly and not applied to conditions other than obtained during the test. 

A knowledge of the water rate of the machine proposed is essential in order that the siae of 
boilers and other parts of the steam plant may be proportioned to generate and handle the rfr* 
quired amount of steam. 

The curves Fig. 25 and Table 11 may be used in this connection. The curves are fair 
averages of the results that may be expected from well designed machines. 

TABLE 11 

APPROXIMATE STEAM OR WATER CONSUMPTION 
Of Various Types of Engines, in PoiindB per Indicated Horsepower per Hour at Nonnal Load 

Non-Condensing 



Hone 

Power 

Rating 

of Engine 


Simple 

Sinfle-Valve 

Throttling 

Engines 


Simple 

Single-Vslve 

Automatie 

Engines 
100 Lb. I.P 


Simple 
Four-Vslve 
Automatic 

Engines 
100U>. I.P. 


Simple 

Coiiiss 

Engines 

100 Lb. I.P. 


Tand^n 

or Cros»* 

Compound 

Pour-Valve 

and Corliss 

Engines 
100 Lb. I.P. 


Tandem 

or Cros»* 

Compound 

Four-Valve 

and Corliss 

Engines 
126 JJb. LP. 


Tandem 
or Cross- 
Compound 
Pour-ValTB 
and Corlifls 

Engines 
ISOLbTLP. 


10 


48 to 62 
46 to 60 
44 to 48 
42 to 46 
40 to 44 
89 to 48 
89 to 43 
88 to 42 
38 to 42 
38 to 42 
37 to 41 
37 to 41 
37 to 41 
37 to 41 
37 to 41 
37 to 41 




86 to 89 
36 to 88 
83 to 86 
31 to 34 
30 to 33 
30 to 88 
29 to 82 
29 to 32 
28 to 81 
28 to 81 
28 to 31 
28 to 81 
28 to 31 
28 to 31 
28 to 31 














20 












80 












40 












60 


27 to 29 
26 to 28 
26 to 27 

25 to 27 

26 to 27 
26 to 27 
26 to 27 
24to26H 
24 to 26 
24 to 26 
23 to 25 
23to24 










60 


26 to 27* 
26 to 27 
26 to 27 
26 to 27 
26 to 27 

24to26H 
24 to 26 
24 to 26 
23 to 25 
23 to 24 






70 


22 to 24 
22 to 24 
21 to 23 
20to22H 
20 to22 
19 to 2m 
19 to 21 H 
19 to 21 
18 to 20 V^ 
18 to 20 
18 to 20 






80 






90 

100 

160 

soo 

250 

300 

400 

600 

600 


i9to2i" 

18to20H 
18 to 20 

18tol9H 

18tol9H 
17 to 19 

17tol8H 
17 to 18 
17 to 18 
16tol7H 
16 to 17 H 


i8":to'26** 

18tol9H 
ITtoW 

17tol8H 
17tol8H 
17 to 18 

16 to 17 


700 






16 to 17 


800 






16 to 16^ 
16 to 16^ 


900 








1000 








16 to 16 




• •••••••■ 


... 









The foregoing table was compiled by the Atlas Engine Co., principally from the records 
of a large nmnber of actual tests made mider favorable conditions. It is, perhs^, not entirely 



STEAM ENGINES 



271 



/ 



/ 



free from individual errars, but is sufficiently accurate to afford an approximate idea of the 
amount oi water, in the form of dry steam, an engine of a certain size and type will require; 
also, a oomprehensive basis for comparison of various types of engines, from the standpoint of 
eeonomical performance. 

When running condensing, with normal load and maintaining 26 inches of vacuum, the 
ooiisiimpti<Mi of steam, or water, is reduced about as follows: With 100 pounds initial pressure, 
25 per cent; with 125 pounds initial pressure, 20 per cent; with 150 pounds initial pressure, 
15 per cent, not considaing the steam used to operate the vacuum pump for the condenser and 
pump the cooling water. 

TABLE 12 

SINGLB-VALVE AND HIGH-SPEED FOUR-VALVE NON-CONDENSING ENGINE ECONOMIES COM- 
PARED AFTER SEVERAL MONTHS' RUN 



Type 

City 

Rnginn run ainee Talves refitted 

;>riMiiire, pounds 

k PI wui 1 

Kw. Of geDerator 

Kw. durins test 

Steam per DOur, pounds 

Lhp. per hour 

Steam per Lhp.-hour 



"Steam- 
Tight'* 
Sinrie- 
Valve 



Boston 

1-2 moB. 

16x16 

103.4 
1.0 

125. 

104.6 
4512. 

178.4 
26.01b. 



"Steam- 
Tight" 
Single- 
Valve 



New York 
6 moa. 
15x16 
98.6 
1.8 
100. 
92. 
4143. 
159.7 
26.96 



"Non- 
Rdeasing 

Gear 
Corliss" 



Mentor, O. 
8 moa. 
14x16 
106.1 
1.0 
100. 
102. 
6166. 
178.6 
28.88 



"Non- 
Rdeasing 

Gear 
Corliss" 



Rochester 
8 moa. 
16x18 
188.0 
1.6 
160. 
143. 
6006. 
223. 
26.9 



Compound 
with Steam- 
Tight Single 
Valves 



Buffalo 

12 mos. 

16 & 27 x 18 

126.0 
1.0 

260. 

236. 
8110. 
883.1 

21.1 



TABLE 13 

RESULTS OF COMPOUND HIGH SPEED NON-CONDENSING ENGINE TESTS 



Siae of engine 

Rated hon»^wer 

Rated capacity of generate, kw. 

Average ateam p ro ss u re, pounds 

Back ple a sure at mid-stroke above atmosplMve, pounds per sq. in. 

Total water fed per hour, poimd 

Loas of steam and water per hour, due to drips, pounds 

Loss of steam and water per hour, due to leakage, pounds 

Moistore in steam at throttle, by uirottling calonmeter, per cent . . 

Net dry steam consumed per hour, pounds 

RevohitioiM per minute 

Avenge indicated horsepower 

HJgb-preasnre cylinder 

Low-pffcawire cyliuder 

^ T otal 

Amjvtndmate load 

Water aa fed, per Lhp. per hour, pounds 

Net dry steam per Lhp. per hour, pounds 



16-1/32 ft 26-1/32" x 20 
320 
200 
132.66 
1. 
6838.75 
87,26 
669. 
1.72 
6004.91 
202. 

146.84 
130.79 



276.13 
18.13 



16H ft 28 X 20 
400 
260 
131.23 
0.96 
7728.83 
97. 
669. 
1.43 
6872.63 
200. 



206.96 
178.89 

385.84 
FuU 
20.03 
17.81 



TYPES AND SELECTION OF ENGINES 

There are many points to be considered in the selection of the prime mover. The most 
economical raigine is the one which will develop a brake horsepower-hour for the lowest cost. 
The cost is made up of several items, each of which has a direct bearing on the subject and are 
as follows : The fixed charges (interest, depreciation, taxes and insurance, the sum of which is 
usually assumed as about 15% of the initial cost), attendance, and suppUes (oil and waste), 
cost of fuel and water. It is evident from the foregoing that the economy, steam or fuel per 
Lhp.-hour, may not under certain conditions be the determining factor where all things are taken 
into account, as, for example, the character of the load and opportunity to use the exhaust steam 
for heating or process work. 



272 POWER PLANTS AND REFRIGERATION 

The plant as a whole, including generators, boUera and auxiliaries, must be oooMenA to 
make a true compariaoD, as the more economical engine requires lesa boiler capacity. The hi^>er 
the engine speed the leas is the cost of geaeratora for direct connected units. (See the Chaptar 
on "Cost of Steam and Gas Power Equipment.") 

When exhaust steam may be utilised for heating and drying purposes am engine giving hi^ 



I 



economy is undesirable when the amount of exhaust that may be utilised is in exccn of the 

amount furnished by the more economical engine or turbine. 

The following are the principal types of engines oSered by various builders in this country: 
High-speed Automatic Single-valve Simple Englnsg. High-speed automatic simple engiDea 

are lAtainable in standard units from 25 lew. operating at a speed of 300 to 400 r.p.m. to 300 kw. 



STEAM ENGINES 



273 



at 150 to 200 r.p.m. These engines are enclosed, self-oiling and equipped with balanced vahre 
and shaft govemcH', which regulates the speed within IHVo from no load to full load. They 
are built in both vertical or horizontal types. Having the fewest parts of any type, they require 



TABLE 14 

DIMENSIONS AMERICAN-BALL SIMPLE ENGINES FOR DIRECT-CONNECTED SERVICE (Fig. 26) 













GufBUL DmsHnoMi m Incbm 




SuiPHHU 
WUOBT 






C^findff 


Rerohitioiu 












nrPouinw 




























Hp. 


Kw. 


Diuneter 
and Stroke 


Bunute 


Floor Spftoe 


Wheel 


C 


D 


E 


F 


H 


Steam and 
Exhaust Pipes 


Direct 
Conn. 


SS- 






























Length 


Width 


Dia. 
A 

48 


w*t 

B 








29H 




Steam 


Rxh't 


Engine 


Dyn. 


40 


25 


/SI z 8 
idx 8 


350to400 


/84 \ 
\86H 


81M 


9 


13 

• 


27H 


161^} 


20H 


I2H 
3 


»^t 


5450 


8500 


10 


S5 


9x10 
10x10 
11 X 10 
12x10 
11 X U 


900to325 


lOOH 

102H 
IIQH 


86H 


54 


11 


14H 


33H 


78H 
TTH 


WH 


34H 


¥ 


4 
4 

4H 


8150 


11700 


» 


SO 


12 X 11 
13x11 
13x13 


276 to 300 


111^ 
121H 


04H 


66 


13 


16 


34H 


78N 
85H 


85H 


27H 


4H 


^ 


10300 


14700 


UO 


76 


14 X 12 
15x12 


260 to 290 


124 


IWH 


72 


15 


18 


39 


88 


38H 


28H 


4H 


5 

5 


14100 


20600 






14x14 




























110 


100 


15 X 14 
10x14 
16x16 


240 to 260 


139 


114?i 


78 


17 


20H 


43M 


lOOH 


42H 


33 




6 


18150 


26600 


SOO 


125 


17 X 16 
18x16 
17x16 


220 to 240 


151H 


125K 


84 


19 


21H 


46H 


109H 


45 


36H 




8 


21450 


32050 


250 


150 


18x16 
18x16 
20x16 


210to230' 


152H 


132^ 


84 


19 


23K 


48 


llOH 


45 


42 


^ 


8 
9 


25450 


88450 




























10 , 






IS 


200 


24x18 


190 to 210 


180H 


147 


90 


23 


26 


54 


136H 


48 


51 


8 


10 


36100 


54000 



V .-» 



J * i. 



TABLE 15 

DIMENSIONS B. F. 8TURTEVANT CO. SIMPLE ENGINES FOR DIRECT-CONNECTED SERVICE (Fig. 27) 



Saaof 
Rngiiui 



8z 
16 z 
16 z 

8z 
10 z 
10 
11 
10 
18 
12 
14 
12 
14 z 
14 z 
16 z 
14 z 
16 z 
16 z 



8 
8 
10 
10 
10 
10 
10 
12 
12 
12 
12 
14 
14 
14 
14 
16 
16 
16 
18zl6 



9tMm 


Revolu- 


PiFBB 


Pressure 


tiona 
M^ute 






Required, 
Pounda 


Steam 


Ezhauat 


80 


876 


2H 


8 


40 


260 


8 


i^ 


86 


260 


8 


120 


860 


1^ 


8 


. w 


860 


4 


120 


800 


8 


8H 


60 


800 


8H 


4 


120 


276 


8 


8H 


80 


276 


4 


^n 


120 


276 


4 


4H 


60 


276 


6 


6 


120 


276 


4 


6 


80 


276 


6 


6 


120 


260 


6 


6 


80 


260 


6 


6 


126 


260 


6 


6 


100 


260 


6 


7 


126 


260 


6 


7 


100 


260 


6 


7 



Crank Pin 
Dia.z 
Length 



4 
4 

4] 

4] 

4] 

4] 

4] 

6 

6 

6 

6 

6 

6 

6 

6 

7 

7 

7 

7 



z 
z 
z 
z 
z 
z 
z 
z 
z 
z 
z 
z 
z 
z 
z 
z 
z 
z 
z 



2! 
2| 
8i 

8: 

8i 

8; 

8^ 

8! 

8! 

8' 
4! 
4] 
4 

4! 
6] 

S| 
6! 

6] 



Shaft. 
Dim. 


Kw. 


8U/h 


20 


8»»/m 


20 


^H 


20 


^H 


80 


^H 


80 


^H 


40 


4H 


40 


6 


60 


6 


60 


6 


60 


6 


60 


6 


76 


6 


76 


6 


100 


6 


100 


7 


126 


7 


126 


7 


160 


7 


160 



Number 

16 cp. 

66 Watt 

Lamped 



866 

866 

866 

660 

660 

780 

780 

910 

910 

1400 

1400 

1.866 

1.366 

1.820 

1.820 

24576 

2.276 

2.780 

2.780 



Weight 
Complete 

Set* 
Pounda 



6.000 

6.000 

8.600 

8.000 

8.800 

9.200 

9.400 

144S60 

16.000 

14.760 

16.200 

19.800 

19.800 

24.000 

24.000 

80.600 

80.760 

88.600 

88.700 



^Carboo filanicat type lamps. Approzimately the sanw number o( 60 watt tungsten lampa rated at borixoatal 
cndto power may be auppUed. 



274 POWER PLANTS AND REFRIGERATION 

a miuiiDUiii of attention, require the least floor apace and are the least expensive to install. 
They are, however, the most uneconomical type in so far as the fuel consumption ia concerned. 




See economy curves for various loads (Fig. 
25). The economy curve is fairly flat frtHn 
H to ^H load, and for H load thb type 
is more economical than the compound 



This type is more often used in the 
smaller sizes up to ISO kw. and to larger 
lations where all i}f the exhaust steam may be 
in locations requiring a compact unit. This 
' used in hotel, office, loft and government 
r plants, and localities where coal is cheap. 
es from SO ta 125 lb. gage are generally used 
with this type. The sices, rating and di- 
this type of engine are given by Titles 13, 
FiQ. 27. SIMPLE HioH-SPBBD Bnoines. ^^ '^^ ^^- ^^"^ ^^^^ ^ "^^ '"* "Baking preliminajy 

(B. F. SlurUtatU Co.). layouts, etc. 

TABLE 16 

DtUENSIONS OF SIMPLE HIGH-SPEED ENGINES C3MPia.IT] 
(B. F. StiuUmKl Co.) 



21 K 



P 



nii 



« ^bt-poi«- 



■re for coavoUvao 



SM6LL-5ISAM 






0/me/73/of7S' 


in /nches 


- /fer 3/ower AHachmen^ 








Size 


A 





c 


£ 


a 


H 


L 


u 


M 





p 


R 


5 


T 


1 


Qsnk 
fin 


CroM- 


1 


*l 


3ltA3 


t4 


Si 


'H 


17 


ei 


II 


10 


10 


■? 


4i 


m 


3k 


3i 


"i 


li 


lll-l\ 


«"< 


/ 


'» 


4 fA 


30i 


II 


2CJI 


/f 


n 


13 


lis 


iii 


13 


3* 


H 


3li 


ei 


«i 


iM 


'i'H 


ih'ii 


li 


It 


4^X4 


^ 


II 


^^ 


/* 


n 


13 


lis 


11% 


13 


3i 


III 


3» 


7i 


36 


li 


21,1-H 


li'H 


li 


li 


S X^ 


3H 


iJi 


^1 


a 


H 


/S 


14 


14 


I'i 


6 


»J 


iH 


Si 


4H 


l* 


!i,3 


li'ii 


li 


2 


6»^ 


Xi 


"i 


25/ 


23 


H 


13 


It 


14 


I'i 


e 


"i 


'^ 


'f 


«* 


^f. 


^'3 


iS"l 


li 


2 


6x6 


4Z 


I6i 


m 


26 


Hi 


ni 


IS 


le 


I4i 


'i 


29 


3i 


"1 


«« 


n 


2a.3i 


i'4'H 


li 


2 


ikxe 


^4 


'H 


^ 


26 


//i 


ni 


IS 


16 


ui 


n 


M 


3% 


"i 


«I 


'i 


!ll-3t 


iS"i 


2 


2i 


7x7 


*7i( 


If 


3'il 


30 


13 


It 


IS 


IB 


^ 


7i 


34 


ei 


10 


331 


'S 


3i" 


'i"i 


2 


2i 


^X7 


*7i 


/f 


VI 


30 


/4 


/7 


le 


18 


n 


n 


34 


7 


Hi 


331 


'X 


3k>'4 


iH'H 


2i 


3 


tfxe 


SZ 


'li 


S6i 


34 


/s 


tz 


m 


!0l 


1 


8 


s 


7i 


13 


(31 


3l 


3llt4f 


2i-3( 


3 


3i 


//A8 


5Z 


zii 


^^ 


34 


/J 


Z2 


lei 


XI 


<I 


a 


SI 


H 


"i 


S3t 


'S, 


311-41 


U'3i 


3i 


S 



Ifze 



Tab/e of/iorse /hiveKs - 7/-pe /?- /^/^ Pressure. 



J/eam Prsssuns np^o^ /^O /£i3. 



2S0\lJS\3dd 



Af//7uf-e 



37J\dOO\425\4X>\475\'S00\SS0\6OO\ WO\BOO 



2.38 ZS 273 30 3S 4X} 
2m 3./Z 344 3.75 433 3J3 



&SP_ 745 



m. 



Theheev/bhck/mesi/ji^^sc^fhe 
3lotresfMe^afi>^/e/>thea[,^3in3s6i 
'em//rvgt/We.^eoff//i«3eivA> 
must be used. 



276 



POWER PLANTS AND REFRIGERATION 



Simple Engines for Driving Fans, Blowers and Centrifugal Pmops. Both vertical and 
horizontal engmes, either direct connected or belted with automatic or throttling governors, are 
used for the purpose indicated. Standard sizes of vertical engines as manufactured by the 
American Blower Co, are given by Fig. 28 as well as* their rated capacity at various initial pice- 
sures and speeds. These engines are usually rated at }^ cut off. 

The limitations in the design of automatic flywheel governors do not permit of the epeeda 
being reduced beyond a certain point as given for the various sizes. The exhaust from the fan 
engine is turned into the first section oi the heater, which is separately trapped and drained. 

The steam consumption of these engines is given by Tabte 17 at their rated loads and speeds, 

TABLE 17 

WATER RATE FOR SMALL HIGH-SPEED AUTOBiATIC ENGINES 



Sise and Speed 


Steam 
Pressure 


Load 




IK 


1 


H 


H 


ZHxZ inehes, 620r.pjii. 


126 

100 

80 

60 


46 
47 H 
48 
49 H 


44 

46 
47 
48H 


46 

48 
49 
60 


61 
64 
66 
66 


4x4 inehee, 600 r.p.m. 


126 

100 

80 

60 


46 
46 
47 
48H 


46 

47 


46 
47 
48 
49 


60 
68 
64 
66 


6x6 inches, 600 r.p.in. 


126 

100 

80 

60 


89 

41 

8« 


21^ 
89 

41 

44 


89 
41 
48 
46 


43 
45 

47 
50 


6x6 inches, 460 r.p.m. 


126 

100 

80 

60 


42 


86 
87 
89 
41 


87 
88 
40 

4m 


41 
42 
44 

46 


7x7 inches, 400 r.p.m. 


126 

100 

80 

60 


87 
89 

41 


86 

87 
89 


86 
87 
88 
40H 


40 
41 
43 
46 


9x8 inches, 876 r.p.m. 


126 

100 

80 

60 


84 
86 


88 
84 
86 
86 


84 
86 
86 

87H 


88 
40 
41 
42 


7 and 12 x 7 inches, 400 r.p.m. 


126 

100 

80 

60 


28 
80 
81 
82H 


26 H 

28H 
29H 
81 


27H 
89H 

81 H 

82H 


31 
83 
85 
87 



For lower speeds the cylinder condensation wiU be greater; which in turn increases the water 
rate and should be allowed for. 

High-speed Four-valve Simple Engines. This type of high-speed engine is equipped with 
four Corliss type valves, with a non-releasing valve gear which permits it being operated at the 
same speeds as the single-valve engine. The four valves give a somewhat better steam dis- 
tribution. 

These engines are enclosed, self-oiling, and equipped with shaft governors which regulate the 
speed within 1 ^i per cent from no load to full load. The economy of this type is somewhat better 
than the simple valve and about the same as the slow-speed Corliss. It is more often considered 
for units of 100 kw. capacity and over and when the cost of fuel is high.. The first cost is 
somewhat greater than the single-valve engine, but considerable less than the slow-speed Corliss 
on account of less weight per horsepower due to its higher rotative speed. The initial steam 
pressure used is cnxlinarily 100 to 120 lb. gage. 



STEAM ENGINES 



Pio. 29. AimucAK-BAii AMOia Ocaaatjsa Bxoim. 

TABLE IS 
ramNSicnra or ahxrican-ball angi£ cohpodmd engines fob direct-connected 





— 









Sbhthm 






SS& 


Rm 


Cbnbsai. DnaNnom m Ihchh 


MPouima 


^ 


iBW 8pM» 


VhMb 






















1 




A 




f 


1 


i 


1 


A 


C 


r 


H 


J 


•K 


H 


P^ 


n 




i 








J 


« 






a 






D 




" 


K 
















a 




















,, 


M 


7fl 




7* 




,.. 






„ 




























!,'!} 




































s« 














































































































S!S 
























ISS 






VB 


19 




1211 


il»K 


MM 








" 


S1,000 





MOOpmni. 
•HffbMt 



poalliim of pUhm t*D nd. 



278 



POWER PLANTS AND REFRIGERATION 



Compound High-speed Automatic Engines. This class of engine is obtainable in the fol- 
lowing types: tandem cross and angle compound. The compound engine is well adi^ted tor 



TABLE 19 

COMPOUND HIGH-SPEED ENGINES 
Non-Condensing 





Engine 


Max. 

Rat. 


Wheds 


Dia. of 
Pipes 


Floor Spacb 










Belted 


. 


Direct Conn. 


KOowatt 


Siaeof 


Bdt 


Capacity 
of 




Dia. 
In. 


PuUey 

Width 

In. 


St'm 
In. 


Ex. 
In. 


Length 
Ft. In. 


Wdth 
Ft. in. 


Length 
Ft. In. 


Width 
Ft. In. 


Dynamo 


7 — 8 and 18 


i 12 


70 
80 
100 
135 
135 
100 
135 
135 
180 
180 
200 
260 
180 
180 
200 
280 
280 
800 
350 
300 
850 


54 
54 
60 
60 
60 
60 
60 
60 
66 
66 
66 
78 
72 
72 
72 
78 
78 
84 
84 
84 
84 


11 

11 

13 

13 

13 

13 

13 

18 

15 

15 

15 

19 

16H 

16H 

16H 

19 

19 

21 

28 

21 

23 


3 
3 

3H 
3H 

4 

3H 
3H 

4 
4H 

5 
5 
4 

4H 

5 

5 

5 

6 

6 

6 

6 


6 


11 


8 
8 
1 
1 
1 
2 
2 
2 
8 
3 
4 
7 
6 
6 
7 
7 
7 

1 
1 
2 


5 
5 
6 
6 
5 
6 
6 
6 
6 
6 
6 
7 
7 
7 
7 
7 
7 
8 
8 
8 
8 


1 
1 
4 

5 
5 
4 
5 
5 
11 
11 
11 
7 
1 
1 
1 
9 
9 
7 
7 
9 
9 


11 8 
11 9 
18 7 
13 8 

18 8 
13 7 
13 8 
13 8 

15 10 

16 10 

15 11 

17 10 

16 10 
15 10 
15 11 

17 10 
17 10 

19 8 
19 9 
19 9 
19 10 


7 

7 

8 

8 

8 

8 

8 

8 

10 

10 

10 

11 

10 

10 

10 

11 

11 

18 

18 

13 

18 


6 
7 
7 
9 
9 
8 
9 
9 
7 
7 
8 
6 
7 
7 
8 
8 
8 

2 
2 


86 — 40 


7 — 8 and 18 


X 12 


6 


11 


40— 60 


9 — lOHandlS 
9 — 10^ and 16 


X 12 


7 

7 

7 

7 

7 

7 

7 

8 

9 

ID 

7 

8 

9 

10 

10 

10 

12 

10 

12 


13 
13 
13 
13 
13 
13 
15 
15 
15 
16 
12 
12 
12 
16 
16 
18 
18 
18 
18 


1 60— 60 


X 12 


75 


10^—12 and 18 
9 —10 Hand 16 
9 —10 Hand 16 


X 12 


75 


X 15 


60— 60 


X 15 


60— 75 


lOH— 12 and 18 
lOH— 12 and 18 
11 — 12 Hand 19 


X 15 


75 


X 16 


100 


X 16 


100 


12 — 13 H and 20 U 


Ix 16 


100—125 


18 —14 H and 22 


X 16 


12&— IGO 


10 U — 12 and 18 


X 18 


100 


11 — 12 Hand 19 


X 18 


100 


12 — 13 H and 20 U 


Ix 18 


100—125 


18 — 14 H «nd 22 


X 18 


150 


IS u — 16 and 23 


X 18 


IGO 


14 — lBHtLnd2A 


X 18 


175—200 


15 — 16 H and 25 


X 18 


200 


14 — 15Huid24 


x20 


175—200 


15 — 16 Hand 25 


x20 


200 









Condensing 





Size of Engine 


Max. 

Rat. 


Wheels 


Dia. of 
Pipes 


Floob Spacb 






Boiled 


Direct C<mn. 


Kilowat 




Dia. 
In. 

54 
54 
60 
60 
60 
60 
60 
60 
66 
66 
66 
78 
72 
72 
72 
78 
78 
84 
84 
84 
84 




Belt 
PuUey 
Width 

In. 


Capacity 
of 




St'm 
In. 


Ex. 
In. 


Length 
Ft. In. 


Width 
Ft. In. 


Length 
Ft. In. 


Width 
Ft. In. 


Dynamo 


7—8 


and 14 
and 14 
$ and 16 
i and 16 
Sand 18 
iand 16 
I and 16 
i and 18 
Sand 18 
and 19 

and 20 H 
and 22 
i and 18 
and 19 

and 20 H 
and 22 
and 23 
and 24 

^and 25 
and 24 

i and 25 


X 12 


70 

80 
100 ' 
135 
135 
100 
135 
135 
180 
180 
200 
260 
180 
180 
200 
280 
280 
300 
300 
350 
350 j 
i 


IL 
11 
13 
13 
13 
13 
13 
13 
15 
15 
15 
19 

16H 
16H 

16 H 

19 

19 

21 

21 

23 

23 


3 
3 

4 

3H 

3H 

4 

4 

4H 
5 
5 
4 

4H 

5 

5 

5 

6 

6 

6 

6 


6 

6 

7 

7 

7 

7 

7 

7 

7 

8 

9 

10 

7 

8 

9 

10 

10 

10 

10 

12 

12 


11 
11 
13 
13 
13 
13 
13 
18 
15 
15 
15 
16 
12 
12 
12 
16 
16 
18 
18 
18 
18 


8 
8 
1 

1 
1 
2 
2 
2 
8 
3 
4 
7 
6 
6 
7 
7 
7 

1 
1 
2 


5 
5 
6 
5 
5 
5 
5 
5 
6 
6 
6 
7 
7 
7 
7 
7 
7 
8 
8 
8 
8 


1 
1 
4 
5 
6 
4 
5 
5 
11 
11 
11 
7 
1 
1 
1 
9 
9 
7 
9 
7 
9 


11 
11 
13 
13 
18 
13 
18 
18 
15 
15 
15 
17 
15 
15 
16 
17 
17 
19 
19 
19 
19 


8 

9 

7 

8 

8 

7 

8 

8 

10 

10 

11 

10 

10 

10 

11 

10 

10 

8 

9 

9 

10 


7 

7 

8 

8 

8 

8 

8 

8 

10 

10 

10 

11 

10 

10 

10 

11 

11 

18 

18 

18 

18 


6 
7 
7 
9 
9 
8 
9 
9 
7 
7 
8 
6 
7 
7 
8 
8 
8 

2 

2 


35 — 40 


7—8 


X 12 


40 — 60 


8—9^ 


X 12 


50 — 60 


8 — 9V 


X 12 


75 


9 — lOV^ 


X 12 


75 


8 — 9V 


X 15 


50 — GO 


8 — 9V 


X 15 


60 — 75 


9 — lOV 


X 15 


75 


9 — 10> 


xl6 


100 


9H— 11 


X 16 


100 


10 H— 12 


;x 16 


100 — 126 


UH— 18 


X 16 


126 — 160 


9 — lOJ^ 


X 18 


100 


9H— 11 


X 18 


100 


10 H— 12 


Ix 18 


100 — 126 


114—13 


X 18 


160 


UH— 13 


X 18 


150 


12H— 14 


X 18 


175 — ^200 


13 —14 ?^ 


X 18 


175 — ^200 


12H— 14 
13 —14^ 


x20 

x20 


200 
200 



most installations in localities where coal is high priced and particularly when the exhaust from 
a simple engine could not all be used advantageously. The rotative speeds are the same as for 



STEAM ENGINES 



279 



flimple highrspeed enginee. The steam pressure used is 120 to 150 lb. gage. A compound en- 
gine is ordinarily used for units of not less than 160 kw. capacity. They are built both with 
single and four valves. The regulation is about the same as for the high-speed simple type. 




Fio. 30. H&iYT Duty Dirbct-Connsctbd Cobliss Engines. 



TABLE 20 
DIBfENSIONS OF SIMPLE DIRECT-CONNECTED CORUSS ENGINES 



SlZB 




List or Gbnbral FouNDAnoN Dimensions 






Diam- 
«tor 


Stroke 


A 


B 


C 


D 


E 


F 


G 


H 


J 


K 


L 


M 

Ft. In. 


N 





P 


Q 


In. 


In. 


PL In. 


Ft In. 


F.L 


F.L 


Ft, In. 


F.L 


F.L 


F.L 


F.L 


Ft. In. 


Ft. In. 


Ft. 


In. 


Ft. In. 


Ft. In. 


Ft.In. 


12 


36 


28 


5 8 


6 


4 10 


6 


2 9 


3 3 


4 6 


14 8 


2 1 


8 


18 


9 


10 3 




7 


2 6 


14 


86 


28 


6 8 


6 


4 


10 


6 


2 9 


3 3 


4 6 


U 6 


2 1 


8 2 


18 


10 


10 6 




7 


2 6 


16 


36 


25 


6 6 


6 


4 


13 


6 6 


2 9 


3 3 


4 9 


15 1 


2 6 


8 9 


19 


7 


11 3 




9 


2 11 


18 


36 


25 


7 


6 6 


6 


13 


6 6 


8 


3 6 


5 


16 7 


2 6 


9 4 


19 


9 


12 




9 


2 11 


20 


36 


26 6 


7 


6 6 


6 


13 


6 6 


3 


3 6 


5 8 


1610 


2 6 


10 7 


20 


4 


12 6 




9 


2 11 


20 


42 


28 


7 


6 6 


5 


13 


6 6 


3 


3 6 


6 6 


17 


2 8 


10 


22 


9 


12 9 


2 


1 


3 8 


22 


42 


28 6 


8 


7 


6 


14 


7 


3 3 


3 9 


6 6 


17 6 


2 8 


10 2 


23 


2 


13 2 


2 


1 


8 8 


24 


42 


30 8 


8 6 


7 6 


5 


16 


8 


3 6 


4 


6 


18 6 


2 10 


10 8 


23 


11 


13 6 


2 


5 


4 


24 


48 


82 6 


8 6 


7 6 


6 


16 


8 


S 6 


4 


6 6 


19 


2 10 


11 2 


26 


2 


14 6 


2 


5 


4 


26 


48 


82 6 


9 


8 


6 


16 


8 


4 


4 


7 


20 


2 10 


12 


26 


2 


15 


2 


5 


4 


28 


48 


33 6 


9 6 


9 


6 


16 


8 


4 6 


4 6 


7 6 


21 


3 


12 8 


26 


9 


16 


2 


8 


4 3 


80 


48 


34 6 


10 


9 


6 6 


18 


9 


4 6 


4 6 


8 


22 6 


3 


13 8 


27 





18 


2 


8 


4 8 


82 


48 


36 


10 4 


9 6 


6 6 


18 


9 


4 6 


5 


8 6 


23 6 


3 2 


14 8 


28 


4 


19 


2 


11 


5 2 



The tandem, cylinders placed end to end and one crank pin, is only built for stationary 
woric as a horizontal machine. The cross compound, in which the cylinders are arranged side 
by side, is built both as a vertical and horizontal engine. The engine has two cranks at right 
angles to one another, placed on either end of the shaft, the flywheel and generator being placed 
on the shaft between the cranks. 



POWER PLANTS AND REFRIGERATION 



^ g 



HI 



SJti'»'"o 



l.lg...ll5S.. 



3gS!SI3l3l5|S|ssp! 



SI5i=S5l5!2|S|ss|«s 



STEAM ENGINES 




Fla. 31. Hm&vt Dun OaooB-OoKPonin) Div*ct-Oohm ■l.thu Ookum BNOima. 



TABLE 22 

DIMENSIONS or HEAVY DUTT CItOSSi«OUPOUND DIRBCT-COMHBCTED CORLISS BNQINSS 



POWER PLANTS AND REFRIGERATION 




FlO. 32. HeAVT DCTT TAKDEU-OOUPOrtm DiBEcr-CoiraxcTED COBUSe BNOtNa. 



TABLE 23 

HEAVY DUTY TANDEM-COUPOTJND DIRECT-CONNECTED 



^ 






i„. 










DU. 


DU. 




1 


1 




































C 1 D 






H 


J 




■■ 




^ 






u 


"■ 


s 






In 






F I 




=•1 




















L 


!! 


20 


30 


?•!■• 




13 Oj6 1 


i 6 




1-^ 




£i 




1 




\' I 


































































































































^t 
























j 




































































































































"1" 






";•• 


'"' 


-' 


3 0« 


lU « 


*'' 


r 


' 


■"1 


• 





STEAM ENGINES 283 

The angle compound (Fig. 29) has a horizontal high-pressure cylmder and a vertical low- 
pressure cylindn*. This engine has only one crankpin. 

The tandem occupies less space, weighs less, and is, therefore, the che^)est of the compound 
group. FiMT these reasons it is perhaps more often installed than the cross compound. 

The angle compound occupies the least floor space and is used principally for isloated plants 
in hotds and oflice buildings where the floor space is very limited. 

Compound engines when operated condensing are 15 to 20 per cent more economical than 
the same type operating non-condensing. Condensing units are not ordinarily installed unless 
the load is above 1000 hp. and the cost of coal is $2.00 per ton or over, as the fixed charges on 
the plant as a whole do not generally warrant the extra expense. 

Corlias Slow-speed T^pe. The Corliss slow or mediiun-speed engine is equipped with the 
rdeasing type of valve gear, which limits the speed to a maximum of about 120 r.p.m. The 
engine, on account of its slow speed, is massive and requires a comparativdy large amoimt of floor 
q)aoe and is highest in first cost. The depreciation is less than the high-speed types, and it will 
undoubtedly maintain its original economy for a longer period than any of the previous types 
described. 

Until the advent and perfection of the steam turbine the Corliss slow-speed engine repre- 
sented the highest standard of perfection reached by the engine builder in this country. It was 
chosen to the practical exclusion of all other types for central station installations. For rope 
driving and belted service in miUs and factories it is still a favorite type. For direct-connected 
units this type is not ordinarily chosen for units of less capacity than about 500 to 750 kw., 
at which point the rotative speed and cost of generators is the same as the four-valve non-releas- 
ing type, the generators being the same size. Approximate dimensions for standard Corliss 
engines are given in Tables 20 to 23. 

The Unaflow Bngine. — ^The latest development in the reciprocating engine field is known as 
the XJnaflow (or Uniflow) engine. The economy curve of this type of engine is remarkably flat; 
for this reason it is particularly well adapted to handle fluctuating loads. 

The unaflow principle has for its object the elimination of initial condensalian, one of the 
greatest losses in reciprocating steam engines. 

With the imaflow engine, the steam enters the cylinder at the ends, after passing through 
steam-jacketed heads; and, after cut off and expansion have taken place, the steam is exhausted 
through ports arranged around the center of the cylinder, which are uncovered by the piston at 
the end of the stroke. The steam has, consequently, a flow in but one direction — Whence the 
derivation {A the phrase " imi-directional flow.'' (Figs. 33 and 37.) 

In the counterflow engine, the steam returns on its path at the end of the stroke, and is 
exhausted at the same end of the cylinder at which it entered. By this method, the cold ex- 
panded steam of considerable volume washes the cylinder walls and head during 50 to 75 per 
cent of the return stroke, thereby cooling them to such an extent that the boiler steam, when 
it is again admitted, is cooled or condensed by coming in contact with the head and clearance 
spaces of the cylinder which have just been cooled by the expanded exhaust steam. 

It is this cooling effect that causes what is termed ** initial condensation," which is very 
much reduced in the unaflow engine, where the ends are kept hot and the center or exhaust belt cool. 

It was to remedy this fundamental defect of the counterflow^ engine that successive expan- 
sion stages were resorted to, as in compound or triple-expansion engines. Superheating has also 
been employed to overcome the above-mentioned difficulty; but superheating cannot be effected 
without some cost in installation and operation, and much of the. apparent gain in the engine 
due to superheating is counteracted by the decreased boiler efficiency. 

Therefore, by avoiding the cooling of all clearance surfaces, in the design of the cylinder 
itself, it is possible to obtain in a single cylinder as many expansions, with a good or better econ- 
omy, as can be obtained in a compound or triple-expansion engine, embod3ring the practical 
feature of a much simpler valve gear, less cylinder and gear lubrication and a higher mechanical 
efficiency. 



284 POWER PLANTS AND REFRIGERATION 

In Ennq>e, the majority ateun plants operate condensing; and eompreaniH) begtna «• 
soon as the piaton covers the central exhaust porta on ita return stroke. ComprMsiMi takes 
place, therefore, during 90 per cent of the stroke, which, even with small olearanocs, doM not cause 
it to rise above the initial pressure so long as the engine is operated condensing, with a fairly 
good vacuum, as shown in the oondmaing indicator diagram. 

If the engine should be (^Miuted non-eondenaing, however, the compression, starting almost 
at the beginning of the return stroke, and with atmospheric pressure instead of vacuum in the 
cylinder, would become so excessive as to be detrimental to the engine, unless large clearanoev 



FlO. 33. SUNNXB Ukatlow BMOim Otumdu. 

were employed. The same excessive and dangerous compresnon would be the result if the vac- 
uum should fail on a condensing unaflow engine. 

To prevent this rise in oompression, if the vacuum should for any reason fail, the Ehiropean 
builders reecnl to eitha uaing snifting or relief valvee, or providing large clearance pockets in 
both cylinder heads, contn^ed by hand-operated or spring-backed relief valves, which relieve 
the cylinder of exceeeive oompression by farcing some of the compressed steam into the cleanuice 
pockets. 

In America, where fully 90 per cent of the steam engines Installed operate non-oondensing, 
it was found that the European unaflow engine would have to be considerably modified. 

The Skinner Engine Co. have adopted the expedient of delaying the oompreaaion by placing 
auxiliary exhaust valves at that point in the unaflow cylinder whra« it is usual to start oom- 
piessiGai in a non-condensing oounterflow engine. 

These valves oome automatically into action as soon as the pressure in the exhaust pipe 
exceeds a predetermined limit. 

The cut (Pig. 34) shows the construction of the auxiliary exhaust valve gear and autcHuatic 
disengaging device, of the Skitmer engine. 

A is the shaft supporting idler B, which is operated by shear cam C. This cam is openiitA 
by the engine valve gear, which is connect«d ia shaft D on the outside of the cam box. 

Wbea the cam C raises the idler B, the latter raises the single-beat exhaust valve, the stem 
of which projects within a short distance of the idler B. The spring around the valve stttn 
has only enough tension to insure quick cloeing when operating at high speeds. The shear cam 
is BO designed that there is practically no sliding action on the idler. Both oam and idler are 
of steel, and are immersed in oil. 



■■■^^ 



STEAM ENGINES 



285 



Fig. Z5 aihowB a leproductioa of an indicator card taken from the Skinner unaflow engine 
fdien operating mm-condenaing. 

Fig. 36 showB the economy curve of a Skinner unaflow engine when operating non-condenaing. 

Fig. 37 shows the cylinder construction of the Nordberg unaflow engine. (Nardberg Mfg, Co,) 

Fig. 38 is a r^roduction of the indicator cards from the Nardberg engine. 

Fig. 39 shows the econcxny curves of the Nordberg engine operated with saturated steam 
at 150-Ib. initial pressure and 150 r.p.nL, condensing with 26" vacuum and non-oondeosing with 
one-half-pound back pressure. 

SELF-CX)NTAINED PLANTS 

A type of self-contained power plant known as the ** Locomobile " is largely used in Europe 
in small and medium-sized isolated plants. This machine is simply a combination of a high- 
grade compound engine mounted on a horizontal fire box tubular type boiW equipped with a 
guperheater and reheater located in the smoke chamber. 

The remarkable low fuel consumption is probably its most pronounced characteristic. Its 
ifiiwlMmiftAl simplicity, small space requirement, ease of supervision and ready access (ot inspeo- 
tion and repairs are advantages of scarcely secondary importance. 

The low fuel consumption may be attributed to the high pressure used superheating and 
rdieating of the steam between the high- and low-pressure cylinder by means of the furnace gases. 
This machine is now built in this country by the Buckeye Engine Company under the trade 
name of " Buckeye-mobile." The exhaust from the low-pressure cylinder is passed through 
a dosed type of straight tube feed-water heater and thence to a condenser when operating 
as a condensing machine. The air, circulating, and feed pumps are belt-driven from the 
engine shaft. 

The fdlowing tests results are reported by the Buckeye Engine Co,: 



TABLE 24 



TMt 


Ratine 


Kw. 


R.P.M. 


StMin 

Ptm- 

Mira 


Initial 

Supor- 

heat 


Low 

PrMBurg 

Superheat 


Ftod 
Water 
Temp. 


Vaoanm 


Steam 

I.up. 
Hour 


Coal 

per 

I.Hp. 

Hour 


Coal 

perKw. 

Hour 


Boiler and 
Super- 
heater 

EiBeieney 


B.t.u. 
in Coal 




86 


98.4 


206 


208 


262 


189 


205 


N.-C. 


a • • • • 


1.856 


2.99 


• a 9 9 


14898 


B... 


88 


96.7 


198 


208 


192 


177 


204 


N.-C. 


12.9 


1.45 


2.88 


74 


14209 


C... 


97 


101.4 


200 


209 


218 


178 


185 


25.7 


9.2 


1.08 


1.80 


76.8 


14209 


D... 


94 


98.4 


201 


202 


297 


219 


140 


25.8 


9.6 


1.85 


2.26 


64.8 
6^.8 


14282 


E... 


94 


98.5 


208 


206 


220 


181 


188 


25.8 


10.41 


1.49 


2.49 


12788 




96 


99.4 


206 


209 


247 


169 


181 


25.6 


9.8 


1.16 


1.94 


76.9 


14186 


C... 


114 


121.8 


209 


208 


282 


178 


182 


24.8 


9.9 


1.196 


1.96 


76.6 


14099 


H... 


186 


146.8 


208 


207 


278 


188 


188 


24.8 


10.2 


1.195 


1.98 


77 


14216 


• • • • • 


108 




248 


210 


171 


66 


192 


N.-C. 


18.8 


1.88 


» • • » 


80.2 


14500 



Vv^ .— yffli-ftWH tffi«f"g atmoq>herie exhaust. 

The capaidtiei and dirnenalona of machinea built by the Biidtoyg EwgiiM Co. are given by Fif. 40. 



FUEL CONSUMPTION IN POWER PLANTS 



The fuel ocmsumption in a steam power plant depends upon the calcific value of the fuel 
used and efficiency of the boilers and generating units. 

The 3rear-round efficiency of a boiler plant may be assumed as equal to approximately 60 
per cent. An efficiency of 75 per cent and over is frequently obtained under test conditions. 
AsBuming an average <A 13,500 B.t.u. for the heat value of the coal, there is available 13,500 X 
0.eO or 8100 B.t.u. per lb. (^ coal burned. 



286 POWER PLANTS AND REFRIGERATION 

Aasuming a boiler pressiu'e 0( 105 lb. gage and 100 lb. at the engiiie throttle and a final 
temperature of the feed wat^r 200°, the generation of one pound of steam at 105 lb. gage 
pressure requires 880 + (341 - 200) or 1021 B.t.u. per lb. Therefore, the geoeration of ixte 



E Gear or Skinneb Bnqikb. 




Fia. 35. Indicator Cabo frou Skinneb IInaflow Enoine — KON-OoNDiNBiiro. 



pound of steam under the conditions above stated requires a fuel consumption of 1021/8100 or 
0.126 lb. The fuel consumption per indicated horsepower per hour based on the water rat«a 
given by the curves Fig. 26 at normal load and the data in the preceding paragraphs ie given 
by the following table: 



STEAM ENGINES 



287 



TABLE 25 

STEAM AND FUEL CONSUMPTION OF NON-CONDENSING ENGINES 



Type of Engiiie 



Sfanpte HIcfa-Speed Smgie-Valve 

Simple Hiib-Speed Foui^Valve 

Compound EOcli-Speed Single-Valve. 
Compound Hlgk-Speed Four-Valve. . 



Steam 



I.Hp.-Hr. 



30 
25 
25 
24 



Kw.-Hr. 



46.5 
38.7 
38.7 
37.2 



Coal 



I.Hp.-Hr. 



8.75 
3.13 
3.18 
3.00 



Kw.-Hr. 



5.81 
4.85 
4.85 
4.65 



To the above figures should be added approximately 4 per cent for the steam required 
for feed pump, plus the amount required for operating any other auxiliaries about the plant, 



40* 






s 



30 



Co 



20 



10 
Ljoad 





















































m 














































































































X 


X 




















































^ 


^ 


















— 


— 






Ste, 


m p tr 






















'^ 


2.9 


R7 


— ^ 


29, 


4$ 


"^ 




30 


1 


Km, 


Hoa 


n 
























































































































































' 






























$ 










19 




Stei m p\ 


r 














I9.i 


8 




■"^ 


■^" 


l^r 


??" 


""■ 


19- 


t6 










9 


.H.l 


"\ 


ur, 




































































































































■ 









































































^4 



'A 



% 



/ 



PiQ. 36. EcoNOMT Curves Unaflow Engine. 
{Skinner Engine Co,), 

19 z 20 Saturated Steam, 136.4 lb. Non-condensing. Atmospheric Exhaust on one-fourth and one-half 

Loads. l}4 Pounds Back Pressure on three-fourths and full Loads. 



to obtain the total estimated fuel consumption. The above figures are bettered by approxi- 
mately 33 per cent for condensing engines. To the resulting figures must, however, be added 
about 10 per cent for operating the condenser auxiliaries and feed pump. 

TYPICAL ENGINE SPECIFICATION 

The following is an extract from a Treasury Department specification for two 100 kw. and 
two 200 kw. direct-connected high-speed engines and generators for the United States Post Office 
and Courthouse, Chicago, HI.: 

Type. The engines to be of the single-cylinder, automatic, horizontal, side or center crank 
type. They shall be designed to operate non-condensing on dry saturated steam at 150 pounds 
gage pressure at the throttle. The speed of the 100-kilowatt generator engines to be not more 



288 POWER PLANTS AND REFRIGERATION 

than 250 revolutions per minute, and the 200 kilowatt generator engines to be not more than 200 
revolutions per minute. ' 

Capacities. The engines to be designed so as to operate most economically wh^i generatofs 
are delivering three-quarter load at the rated voltage and speed, and shall be capable of operating 
the generators for two hours when delivering 25 per cent overload at rated voltage. 

Foundation. Foundations to be of the required form to suit engine and generator sub-baaesy 
to be constructed of 1 : 2 : 3 concrete, with the bottom not less than 5 feet below the floor line. 
The top must extend not less than 6 inches beyond the edge of sub-base frames all around and the 
batter in the depth specified to be not less than 3H ^^t each side. Concrete foundation to be 
provided with cushion of 6-inch deep sand. Foundation bolts to be provided with washers and 
wrought-iron sleeves. 

Sub4>ase8. Each engine to be provided with a heavy and substantial cast-iron Bub-base, 
upon which shall be mounted the engine, the sub-bases of the engines to be extended under cylindefs 
for support of cylinders. 

The sub4>ase of generators must be secured to sub-base of engine in a suitable manner and 
both sub-bases to be secured to the foundations. 

Frames. Each engine to be provided with a heavy and substantial cast-iron fraiAe deaigDed 
for strength, rigidity, and compactness, and be provided with suitable covers to prevoit throwing 
of oil and allqwing dust to come in contact with the moving parts. 

Bearings. Bearings shall be long, well proportioned, and dust proof. The main bearing 
to be of the removable-shell type and the outboard bearing to be of the oil-ring type. Bearings 
to be lined with genuine babbitt metal carefully peened in place and accurately bored to gage. 
The outboard bearings to be provided with large-sixe oil weUs, visual gages, and pet cocks for 
drawing the oil. Bearings to be provided with means for adjustment. 

Lubricating System. Each engine to be provided with an automatic self-lubricating con- 
tinuous circulating system which shall supply pure, clean oil continuously to all bearings, etc, 
the operation of system to be positive and free from throwing or spilling the oil. 

Cylinders. Each cylinder to i)e made of best grade of close-grained cast iron, bored true 
and smooth, and of sufficient thickness to allow for reboring. The cylinder to be weU lagged 
with magnesia or other material having equal heat insulating value and covered with ornamental 
cast-iron jackets or with Russia iron, properly secured to the cylinder casting. 

Pistons. The piston heads shall be hollow cast iron, with at least two sni^ rings with lap 
joints, made from first quality of hard, close-grain cast iron sprung into accurately fitting grooves. 
Rings shall override the bore of cylinder. Piston rods to be best quality nickel steel. Rods to 
be turned to a taper at the piston ends and each driven up to a shoulder and be securely held 
by a heavy nut to be drilled and provided with cotter or dowel pin. The forward ends to be 
screwed into crossheads and provided with a jam nut and suitable lock to prevent turning. 

Crossheads. The crossheads to be made of cast steel, and be provided with adjustable 
bronze shoes circular in form; shoes constructed of cast iron and babbitt will be acceptable. 
Crosshead pin to be made of steel hardened and ground and held in place by tfq)er fit 
and nut. 

Connecting Rods. The connecting rods to be forged open-hearth steel in one {Meoe with solid 
crank-pin end and crosshead end. The crosshead boxes to be made of phosphor brcnue adjustable 
by means of wedge. Crank ends to be fitted with boxes of steel or phosphor bronxe and lined 
with genuine babbitt metal peened and bored to fit the pins and be adjustable. 

Crank Shafts. Crank shafts to be constructed of open-hearth steel forged in one piece, 
with counterbalancing crank discs of annealed steel, securely fastened thereon. 

Valves. Each engine to be fitted with four valves of the semi-rotary, poppet, or gridiron 
type, designed to be slightly unbalanced and securing positive steam-tight seating over the 
admission and exhaust ports. Steam valves to be of the multiported type, giving ample port 
openings for all points of cut off. Exhaust valves to be designed to give ample port area and 
insure tightness. All valves to be constructed of best quality hard close-grain cast iron. Hie 



STEAM ENGINES 289 

gteam valves to be provided with removable buehinga or c^es; gridircm valves to be provided 
irith suitable balancing plate. 

Vatvfl Mechanism. The valve mechanism on each engine to be desigoed to give quick 



Vw. 87. NosDBBBQ Ukiflow Snoinb OrLDnxs. 



Bcllir Prtmn 




FlO. 3S. IKDICATOB GaBDS nOU NOBDBIBa DHIFLOW ENOUIS. 



coNDEMsnra. 

B NOBDBBBO USIFLOW ENOINI. 

and positive motion to the valvee In opening and closing. All pins subject to wear to be made 
of steel, hardened and ground. Alt boxes for pins to be made of phosphor bronie and to be 
■lijustable without filing; boxee constructed of steel with bronze bushings will be acceptable. 
LubricatioD of pins and bearings to be accomplished while in motion by compression grease 
■raps placed at aoceaaible points or to operate in oil wells. 



290 



POWER PLANTS AND REFRIGERATION 



Bccentrics. The eccentrics to be strong and Jight to reduce the strain upon the governor 
springs. The eccentric straps to be lined with best quality of antifriction metaL Ample 
of lubrication to be provided and designed to be free from oil throwing when in motion. 

Governors. Each engine to be equipped with an inertia governor of approved type. 

The governor pins to be made of steel, hardened and groimd true. The lever-arm 
to be made with hardened steel bushings and rollers. 

Steam Consumption. Each bidder must state in his proposal sheet the iqieed, indicated 
horsepower, full load of each engine. 

The minimum steam consumption when operating under conditions herein specified at 
uniform loads must be stated. 

Each engine when operated under conditions herein specified and at uniform loads most 
not consume more than amoimts of dry steam in pounds per kilowatt hour, determined by the 
weight of condensed exhaust steam for each load as stated below: 









Load 








25 per Cent 


50 per Cent 


76 per Cent 


100 per Cent 


125 per Cent 


100 kilowatt generator engine, dry 
■team 


74 
74 


48 
45 


41 
40 


41 
40 


48 


200 kilowatt generator engine, dry 
■team 


41 







Shop Test of Engines and Generators. The efficiency, capacity, etc., of each unit to be 
determined by actual test in the presence of the department's authorized agent, who shall deter- 
mine the test conditions. 

The tests are to be made at the shop where engines are constructed, and to begin 10 days 
after receipt of notice from contractors of their readiness to commence tests, and to be at Uie 
expense of the contractors, except traveling and other expense of the department's agent. The 
generators to be shipped to engine builder's shop for test. 

Each unit to be run at one-fourth, one-half, three-fourths, full, and one and one-fourth loads 
for one hour under each load, during which time the exhaust steam will be condensed and weired 
and indicator cards taken as often as deemed necessary. 

Engines to be run at the speeds specified with steam at 150 poimds pressure per square inch 
at the throttle, quality of which will be determined by throttling calorimeter placed in steam 
pipe above throttle. 

Penalty. It must be distinctly understood to be one of the conditions under which bids 
are submitted for the woric embraced in the specification that the engines and generators will 
meet every requirement of the specifications and the guaranteed amounts for steam consumption 
named by bidder, under which conditions the contract price will be paid. In event the units 
fail to meet the specification requirements or the steam consumption is greater than that guar- 
anteed by the bidder, the department shall have the right to reject the unit or units abediutely 
and require the supply of satisfactory unit or unite which shall comply with all contract require- 
ments in regard thereto; or if it elects to accept same in event steam consumption, at any load, 
is greater irrespective of other loads than that named in the proposal, then the contract price shall 
be the amount named in the contract for a satisfactory plant less the amoimt of deficiencies shown 
by test based on the following schedule for each pound or fractional part of a pound of steam 
per kilowatt hour: 





Load 




25 per Cent 


50 per Cent 


75 per Cent 


100 per Cent 


126 per Cent 


100 kilowatt-hour unit 


$30 


1120 
240 


1420 
840 


$180 
360 


$75 


200 Idlowatt-hour unit 


60 


160 







STEAM ENGINES 291 

Plmt Test. At expitation of three monthB' operatiiig test, b test will be made to deteimine 
Uk Bteam oonsumptton per kilowatt-hour under (q>erBtiiig conditions and it will tarm a bans f(» 
compariaixi of steam oonsumption at expiration of one year from that date. 



IL 


; 1 S 




ii 




1 




a 




1 



!i 



re not to be reground or sonped during & period of one yvai and steam consumption 
e year must not exceed the amount ascertained by steam met«T as above not«d. 



292 



POWER PLANTS AND REFRIGERATION 



Steam oonsumption to be measured by recording steam-flow meter now installed <m the premiseB 
for both tests. This instrument will be calibrated before each test. 

In the event it is found that the steam consumption is greater at end of year, ihea the de- 
partment reserves the right to require the contractor to make such changes in the engines as it 
elects at the expense of the contractor. 

The supervising architect reserves the right to waive these tests or any portion thereof, 
and require contractor to submit certified test sheets, in triplicate, for approval, it being under- 
stood that those portions not waived shall be exacted when apparatus are installed if not per- 
formed at shop as specified above. 

Regulation. After engines are installed in position they must be adjusted to run smoothly 
and practically noiselessly. They must be tested at shops for regulation, which tests must 
show that slow change of speed from no load to full load and vice versa will not show more than 
lyi per cent variation and from full load suddenly thrown on or off the variation shall not exceed 
2 percent. 

Fitthigs. Each engine to be furnished with the following fittings: 

One throttle valve. 

Automatic cylinder relief and drain valves. 
Mechanical cylinder lubricator, piping, etc. 

Metal packing (approved) for piston rods and valve-stem stuffing boxes. 
Auxiliary hand oil pump. 
Steam-chest drain connections with valves. 
Indicator piping with three-way cocks and angle globe valves. 
Attached indicator reducing motion. 
Set of adjusting wrenches on hardwood or cast-iron board. 

All necessary drip, drain, and indicator piping, which must be brass, nickel plated, 
exposed above flow. 

Paintingi Etc Engines and generators to be filled, rubbed down, and after installed in 
building to be finished with two coats of paint, color to be very dark green, then striped with gold 
leaf and varnished. 



TABLE 26 

SMALL RECIPROCATING ENGINE DRIVEN GENERATOR SETS. 

APPROXIMATE SELLING PRICES 



RELATIVE WEIGHTS AND 









Eff. 


Eff. 


R.P.M. 


Wbght 






Pmcb 




Type 


Gen. 
Kw. 


E^^ne 


Hp. 


Hp. 




















Gener- 


Com- 




Geneiw 


CoB»» 








Off 


Off 


Eng. 


Gen. 


Engine 


ator 


plete 


RnirliMi 


ator 








Direet-eonnected 60] 
























cyde, 2 or 8 phase 


110 


15"xl4" 


160 


198 


277 


18,600 


10,800 


24,400 


$1,508 


$1,621 


$8,029 


' A.C.generator.Aiiy 
voltaie up to 2200. 


126 


16"xl6" 


196 


283 


277 


21.000 


10,800 


81,800 


1.849 


1.621 


8,870 


126 


16"xl6" 


196 


288 


226 


21,000 


12,800 


88.800 


1,849 


1,980 


8,829 


Ezdternotind'd'd. 1 
























Direet«oimeeted D.' 
C. generator 125 or 
240 volte Com- 


100 


14"xl4" 


162 


194 


290 


18,100 


11,800 


24,900 


1.418 


i^ 


2,660 


126 


16"xl4" 


181 


206 


260 


18,900 


14,800 


28,700 


1.697 


1,666 


8,168 


126 


16"xl6" 


196 


288 


210 


21,000 


18,000 


89,000 


1.849 


2.084 


8,888 


pound wound 


186 


16"xl6" 


196 


283 


226 


21,000 


16,000 


87,000 


1,849 


1,710 


8,669 


Belted, 60 cyde, 2 or 
























8 phase A. C. gen- 


100 


14"xl4" 


160 


174 


275 


900 


12.000 


8,100 


20,100 


1,188 


1,071 


2,259 


erator. Anv Tolt- 
ageupto2200. Ex- 


116 


14"xl4" 


160 


174 


275 


900 


12,000 


8,100 


20.100 


1.188 


1.071 


2,250 


160 


16"xl6" 


210 


248 


260 


900 


19,000 


9,800 


28.800 


1,676 


1,886 


2,961 


dter not induded. 


























Bdted, D. C. geueia- 
tor 126 or 240 volU 


100 


16"xl4" 


183 


220 


275 


660 


12,750 


9,860 


22,100 


1,868 


1,184 


2.502 


100 


16"xl6" 


196 


288 


226 


460 


19,000 


18,000 


82.000 


1,676 


1,486 


8,060 


Compound wound. 


126 


16"xl6" 


210 


248 


250 


660 


19,000 


18,000 


82,000 


1,676 


1,486 


8,060 



Nan.— Only dmple, automatic, non-eondendng enginee are listed above. 



ffTEAM ENGINES 



W///M///M- 




f/M/M/M/M Vm/M/MMA • V7777P7^y 



yj/7///////j/j'/j/A v;/j/^//////j///M' ' ' ~r^f^////^ 



ho. 41. Dbcui. Layout iok Two <w Mob> Bxnxmim-MoBiut UNm. 



294 



POWER PLANTS AND REFRIGERATION 



TABLE 27 

APPROXIliCATE PRICES OF HIGH-SPEED 4-VALVE ENGINES 



F.03. Worlo-^Cort of Genenton Indoded in Priee 
Simple i-Vahre Horiiontal Type 






Hp. RatiBK 


Stroke, Indiee 


Siseof 

Generator 

Kw. 


R^.M 


Net Priee 


160 


16-18 
18-20 
24-27 
24-27 
24-27 
24-27 
80-«6 
80-«6 


100 
160 
200 
260 
260 
800 
400 
460 


286-250 
215-285 
160-200 
175-200 
160-166 
160-200 
140-160 
125-160 


12.960 
8.400 
4300 
4,700 
6,400 
6,660 
7460 
7,600 




225 




800 




875 




875 




450 




600 




675 




Tandem Horiiontal— ^-Velve Type 


160 


16-18 
18-20 
24-27 
24-27 
24-27 
24-27 
80-86 
80-86 


100 
150 
200 
260 
260 
800 
400 
460 


286-260 
215-285 
160-200 
176-200 
160-165 
160-200 
140-160 
126-160 


18400 
4,500 
6,700 
6400 
7460 
7.750 
9,700 

10,000 




225 




800 




875 




875 




460 




600 




676 










Croes-Compound — 4-VBlTe Type 






260 


16-18 
16-18 
18-20 
18-20 
24-27 
24-27 
24-27 
24-27 
80-«6 
80-86 


160 
200 
260 
800 
800 
400 
600 
600 
760 
900 


200-226 
200-225 
176-226 
200-226 
160-176 
160-200 
160-176 
160-175 
160-160 
126-160 


85,600 

5,950 

6,600 

6,800 

7.550 

7450 

10,000 

10,400 

12.750 

18400 




800 




450 




460 




600 




760 




900 




1,125 




1.860 



CHAPTER XI 

STEAM TURBINES 

Tlie steam turfoine, owing to its compactness due to its high speed and the absence of many 
moving parts, is rapidly r^lacing the reciprocating engine for electric service, especially when 
condensing installations are considered. It has become the standard for central station work. 
Hie economy is largely dependent upon the efficiency of the condensing apparatus employed as 
the turbine depends upon a high vacuum being maintained to show its best economy. The 
economy fcH* the same degree of vacuum, 24" to 26", is about the same as a high-grade recip- 
rocating engine of equal capacity. 

The turbine, howev^, is capable of successfully operating at 28" to 29" of vacuimi with 
a correqx>nding increase in economy whereas the reciprocating engine, owing to limitations in the 
design oi the exhaust ports and passages, is not capable of handling the large volume of low-pres- 
sure steam generated by extremely high vacuums, without a corresponding increase in back 
pressure due to frictional resistance of the small ports. This tends to offset the gain due to the 
increased vacuunu 

Am<Hig other advantages of the turbine may also be mentioned the close speed regulation 
and the fact that high-speed machinery can be driven directly without the expense or danger in- 
cidental to belts and ropes. The shaft can be placed horizontally or vertically, according to the 
requirem^its of the driven machine. The uniform and continuous flow of steam to the turbine 
permits oi smaller and less expensive steam lines, which in turn reduces the loss of heat by 
radiation. 

Of especial importance in relation to exhaust steam heaUng is the fact that the exhaust is not 
polluted by cylinder oil, also that the theoretical economies of superheated steam can be realized 
without introducing lubrication difficulties. 

Most auxiliary apparatus installed in steam power plants is adaptable for direct driving by 
steam turbines; particularly so are centrifugal boiler-feed pumps, circulating pimups, hot-well 
pumps, centrifugal blowers and compressors, exciter dynamos, etc. Centrifugal high-vacuum 
air i)umpe are also coming into use and by means of gears the tiu*bine can be adapted to driving 
stow-speed induced draft fans, coal and ash conveyors, reciprocating air pumps, automatic stokers, 
large low-head circulating pumps, etc. 

The economy of non-condensing turbines is considerably below that of the reciprocating 
engine and unless the demand for exhaust steam for heatiag or process work is sufficient to con- 
sume practicaUy all of it the high-speed non-condensing engine is ordinarily preferred. The 
turbine will maintain its original economy over a long period. The reciprocating engine depends 
ap(Hi the tightnesfl of the valves and piston for economy and unless careful attention is given to 
keeping them ti^t the economy, after a period, may become no better than a turbine of equal 
capacity. The economy of a turbine is, however, more seriously effected by high back pressure and 
moist steam than the reciprocating engine and these facts should be borne in mind when consider- 
ing the installation of a non-condensing turbine. 

The steam consumption of non-condensing turbines from 35 to 300 kw. at 100 lb. gage 
initial pressure and atmospheric exhaust is given by Table 1, as well as the corrections to apply for 
other conditions that may obtain. 

Corrections for Change In Operating Conditions of Steam Turbines. In order to obtain the 
water rate of a steam turbine (H)erating under conditions as to pressure, superiieat and vacuum 
other than as reported, the following corrections may be ap))lied. 

295 



296 



POWER PLANTS AND REFRIGERATION 



TABLE 1 

WATER RATE CURTIS NON-CONDENSINO TURBINES 
100 Lb. Gagb Initial FuamuBm — ^AncoePHBRic Exhaust 





Stbam Conbumption; 
Lb. pbb Kw^Houb 


Rated Ctpadty Kw. 


Lowl 




H 


1 


IK 


86 


86 

72.5 

60.6 

68 

67 


68 

69.5 

49.6 

66 

68 


68 


75 


67.6 


100 


48.6 


125 


66 


800 


62.6 







NoTB. — Deduct approzimately 2% from valuei sfven in table for every 10 lb. inereaae in tbe initial 
Add 2% for each 1% of moisture m ateam. Inereaae water rate by the following amounta for bade 
2 lb., 8%; 8 lb.. 5%; 4 lb., 7H%: 6 lb., 10H%. 

Superheat, Decrease water rate 1 per cent for each inereaae of 10® F. superheat f(v 0^-100^ 
superheat and 1 per cent decrease for each 12*^ increase in superiieat frcHn 100^-200'* superheat. 

Moisture. Increase water rate 2 per cent for each 1 per cent increase in moisture. 

Pressure, Decrease water rate 2 per cent for each 10 per cent increase in initial pressure 
between 100 and 180 lb. gage pressure and 1.9 per cent for pressures 180 to 200 lb. gage. For 
low pressure turbines decrease water rate by 4 per cent for each 10 per cent increase in initial 
pressure. 

Vacuum. For increase in vacuum from 26" to 27" decrease water rate 5 per cent. For in- 
crease in vacuum from 27" to 28" decrease water rate 6 per cent. 

For increase in vacuum from 18" to 28 H" decrease water rate by 3.87 per cent. 

For increase in vacuum from 28)^" to 29" decrease water rate by 6.75 per cent. • 

For low-pressure turbines the decrease in water rate is approximately as follows: 



// 



12 % for increase in vacuum from 26" to 27" 
'^^%% for increase in vacumn from 27" to 28" 
8K% foi" increase in vacumn from 28" to 28}j 
11 Ji% for increase in vacuum from 28 J^" to 29" 

The expected water rate (IF R) (or a change in condition may also be calculated by multiplying 

the test water rate (IF Rt) by the ratio of the water rate of the Rankine engine for new oonditioD 

(Wn) to the water rate for test condition (TF|) also based on the Rankine engine (page 266). 

Ei 
otW R ^W RtXr^ 

En 

This method of correcting the actual test results for the conditions as specified by a guarantee 
is also applied to steam engine tests within limits. 

Elementary Theory. A steam turbine may be defined as a machine designed to utilise tbe 
energy of steam flow for mechanical work, the force required to retard the weight of nq>idly 
moving vapor being applied to buckets or blades attached to the periphery of a rotating disc or 
drum. From mechanics the change in kinetic energy of a moving mass of W pounds having 
an initial velocity of u^, and final velocity of wt ft. per sec., is: 



2g 



ft.-lb. 



Let W = the weight of steam issuing from a nozzle, lb. per sec. If the stream or jet having 
a velocity of Wi ft. per sec. be directed on a turbine blade or bucket having the shape or form aa 



STEAM TURBINES 297 

Aown by Pig. 4 «ad the blade is held Btationary the stream will issue from the blade with the 



Braka Horsa Power 



wthato 
B Rata CortiH" ibow tb* vudatlon la ntar, or mora comcUy, in ■team, BouuiapClon par luns- 
iukisdi,La.,t£a "Witcr or StawnHmla"a(llieturbliia. £>dl curve eomapoadi to ■" Wmtar 
s- Une, ths lower to the lowcc liMt Mt- 
OiantlTi «adltl<— oovand by Clia taiti «n: 
- ■ ^ ' ' ^ gj^ luperbsatdd ktflun. 

WNoo 

nC Dvarload- 

id teeU, the optrmti/^ of the autommtle BBftnidary nlve mmy be obecrred. It eomee Into eedon 

ti kad u indleUed br ■ bend ia the nter line. With the »ld ol tUi vain the b«(t economy 

It the rsKf* ol nonn*! luedini, while iufo overioid npuity ti aniUbie when de- 

tt rii^tlr dan— d aflldBicy. Wlwn the Hcondary Taive, howerBr, liai some lairly into utian..tha 

■■ indaal hnprorament. ai ibowa by the rercnal o( eumture at tlw Water Rata eurvei. 

V ecomnny iHth auparbeatad ataam at variaua loadi lor both eondenainc 

ta ia wall ihown by the distance betwsen the two pain ol Water Rata eurna. 

work wHI not operate aon-andeiiiliic with quit* M food eeaDOCii]i *■ it daricBed 
. That thia economy K bowenr, eieellent, la ahown by tlM upper pair of enrTca. 
K rat* la aomewhat lea than double the eondenilDi water rata. 

If, however, the blade is moving with a velocity of c ft. per sec. the stream will leave the 
blade with an aboolute velocity, trt — ici — 2c. 




POWER PLANTS AND REFRIGERATION 






STEAM TURBINES 



Tbe eoBTBy impArtod to the bfade la equal to K, equation (1), ft.-lb. per see. 

It the bucketa on moving witb a velooity c •■ K toi, tlien the eoergy absMbed by tl 




to too 200 300 400 too tOO 700 800 BOO 1000 

AbsoluU Pnssan: Lb. Pr. Square Ineh, 



diod tlw work dcMie a 



Pin. 4. 

In this oaae w, -w,-2e-un-2xHtBi-0 



300 POWER PLANTS AND REFRIGERATION 

Impulse and Reaction. The magnitude of the force exerted by the jet on the blade may 
be determined by reference to the formula from mechanics: 

F(»ce "- mass X aocelaration, 

F - — X a lb (2) 

9 

in which " a " is the negative acceleration of the jet or stream in the direction of motiim of 
the moving blades. The impulse effect of the jet on the form of bucket. Fig. 4, is: 

W 
Fp'' — (tci - c) lb. 


As the jet is turned through an angle of 180 degs. by the bucket the reaction produced oti the 
bucket is equal to the impulae so that the total force acting on the bucket is equal to 

j,,2W(u,-c)^ ^3j 



The work performed is: 



2cW(wi-c) . 

Fc ^ ft.-lb. per sec C4) 



Available Energy qf Steam, The energy of steam may exist in two forms, heat and kinetic 
energy. When confined, as in a boiler» under pressure, the energy exists wholly in the form of 
heat. 

Its capacity for doing work lb analogous to that of water stored in a reservoir. 

The available energy of the water is dependent upon the head or diff^^enoe in elevation 
of the water in the reservoir and some lower level at which a water-wheel may be located. 

In the case of steam under pressiu^ the available energy is dependent upon the difference in 
pressure in the boiler and the pressure maintained in the vessel into which the steam is permitted 
to flow. 

According to the Law of Conservation of Ekiergy, the total energy in a pound of steam during 
expansion from a higher pressure (pi) to a lower pressure (pi) remains constant. The potential 
energy of one pound of steam is: 778 ti ft.-lb., in which 778 is the mechanical equivalent of heat 
and ti the heat content of one pound above 32*^ F. at an absolute pressure of pi. The 
energy of one pound of moving steam at a lower pressing, pt, is equal to the remaining potential 

energy (778 ti) plus the kinetic energy—- or 778 it + -r— . In which %% is the heat content 

2g 2g 

corresponding to pressure pt and wt the velocity in ft. per sec. Expressing the law in the form 
of an equation, we have: 

778 t'l - 778 ti + ^ 

2g 



or Wt « 224 V ti — ti ft. per sec. (5) 

This is the theoretical velocity attained at exit from a properly constructed noisle. 

The above equation, when applied to straight or converging tubes, holds for all differences 
of pressure so long as the lower absolute pressure does not exceed the critical pressure. The 
critical pressure or the lower absolute pressure pt, above which the velocity does not further 
increase, is equal to 0.58pi as proven both by theory and experiment. This ratio 0.68 inyplies 
to dry saturated steam. If the steam is hi^^ily superheated initially the ratio is about 0.55. 



STEAM TtTRfilNES 



THE IMPUIBE TURBINE 



In a tuibine of the pure impulse type the expAiuioa and poative acceleration of the ateam 
take place ooly in stationary noizlee or guide vane paBsagea. If the complete expansion takes 
place in a single aet of noulea and the jet is directed on a single wheel or ntUx the turbine is 



Fra. S. Tarn Uatn BLmmna or tbb d^ Lital Tvbbdi 



no. flt JUDUXomntT or Snui Noolb wtih Tuhbinb Whml or Db l^vu, TtrsBniB. 

tenned a littgh-Tmtnm lirtf/le-velodly ttaf/e machine, the preMu« being the same on both sides 
of theirtieeL 

The first commercially Bucceesful machine of this type was the De Laad turbine. The ar- 
rangttneDt oi steam nouleB with turbine wheel ia shown by Fi^ S and 6. The construction eS 
the wheel and blades for a 20 hp. De Laval turbine ia given by Figs. 7, 8 and 9. 

Hie dimensions are ^ven in miliiffletera. The velocity of the ateam jet in the single pressure 
stage turbine variee approximately from 2S00 ft. per sec. for non-cmidensing to 3800 ft. per see 
for ecxtdensing tq^eration. In order to absorb the kinetic energy of steam flowing at auch high 
veloeitiee with a single wbeel or row of blades the peripheral velocity is necessarily very high, 
peripheral veloeitiee ranging from 500 to 1300 ft. per aeo. and whed speeds d 10,000 to 
30,000 r.p.m. bong oommoa. In orda to utilise auch high speeds in practice the wheel or rotor 
qjeed is reduoed by suitable gearing as shown by Fig. 10. The gear ratio is made approximately 
ItolO. 

The ^leed of the A: Laaai turbiiw ia otmtroUed by means of a throttling governor. 



302 



POWER PLANTS AND REFRIGERATION 



The following table gives the speeds and siae of wheels that have been employed in this t3rpe 
of turbine: 

TABLE 2 



Sixe of Turbine, d.hp. 



6 

16 

80 

60 

100 

800 



DianuWlieel 

B«low Blades, 

Inches 




Rer. per 
Minute 



80000 
24000 
20000 
16400 
18000 
10600 



i^Mral 
Vel..>trper 
See.'^C'^ 



615 
617 
774 
846 
1115 
1878 



Steam Nozzles, In order to determine the dimensions of steam nozzles required for a 
particular turbine an estimate of the weight (W) of steam that will be required per sec. must first 
be made. This weight may be approximated from previous tests made on similar turbines. See 
water rate curves, Figs, h 2 and 3, and corrections to apply for other oonditimis of operation: 

Let Pi ^ absolute initial pressure at entrance to nozzle. 
Pp B absolute pressure at throat of nozzle. 

« 0.68pi. 
pt " absolute terminal pressure. 

»i B initial heat content corresponding to pressure pi, B.t.u. 
a r + 9 dry and saturated steam. 
" xr + q wet steam with quality x. 
" r + q + Cp (jti—ts) superheated steam. 
%a "= heat content corresponding to pressure at throat p^ after adiabatic expansion 

from pi to Pm. 
is «B heat content corresponding to terminal pressure ps after adiabatic expansion 

from pi to Ps. 
W a weight oi steam flowing per sec., lb. 

» Estimated water rate for turbine per brake horsepower-hour X brake horsepower 
' divided by the no. nozzles X 3600. 
Am » Area of nozzle at throat, sq. ft. 
Ag = Area oi nozzle at exit, sq. ft. 
Wm *= velocity of steam at throat, ft. per sec. 
Wg s theoretical vel. of steam at exit, ft. per sec. 
Vm » specific volume <^ steam corresponding to p^. 

™ ^m ^"m (approx.) v^'fn a sp. volume sat'd steam p^, 
Vt a specific vol corresponding to ps. 

B xst^^'s (approx.) v"$ a sp. volume sat'd steam ps. 

X«=^ (6) 

Ag^ sq. ft (7) 

ti;„ = 224V ti -!„ (8) 

w. -224V»i-ts (9) 

The values of i^ and t's may be readily found by making use of the entropy tables. 
The MoUier Chart Problems involving the adiabatic expansion oi steam are most con- 
veniently and rapidly solved by means of the MoUier chart. Fig. 11. 



STEAM TURBINES 



303 



An eammple showing the method oi using the chart f oUows. 

Enmple. Calculate the throat and exit diameters required for each of 4 nossles to be used in a 
single pressure stage turbine of 100 brake horsepower capacity. Initial pressure below governor valve 
(ling pressure) pi ^ 140 lb. per sq. in. absolute; terminal pressure pt » 15.7 lb. per sq. in. absolute. 
Estimated water rate of turbine, 38 lb. per d.hp.-hour. Initial condition of steam dry and saturated 

38 X*100 
(xi >- 1.0) W « ^— — B 0.264 lb. weight of steam flowing through each nozzle per sec. 

4 X 3600 

Pm *" 0.58 X 140 B 81.2 lb. per sq. in. absolute. Locate the initial condition of the steam at the 
mtersectlon of the 140 lb. pressure Une and the saturation curve. The heat content as read on the left- 
hand vertical scale is: t'l «* 1194. From the intersection on the saturation curve, above noted, pass 
▼ertioally downward to the intersection with the diagonal pressure line corresponding to Pm » 81.2. 
The quality Xm is found to be 0.96 and the heat content 1150. Ck>ntinue down on the same vertical 




Fig. 7. Wheel Disc. 



until the intersection with a diagonal pressure line corresponding to ps » 15.7 is reached. The quality 
xt is found to be 0.878 and the heat content 1035. From the steam t ables v"m " 5 .41 and v^t -■ 25.22. 
9 ^ ^ 0.96 X 5.41 - 5.19. Ui » 0.878 X 25.22 - 22.2. tr,» » 224 V 1194 - 1150 - 1485. tc, - 224 
V 1194 — 1035 B 2824. The velocities w^ and ir« may be read direct by means of the combination 
B.t.u. and velocity scale provided with the chart. The actual noszle exit velocity wi is less than the 
theoretical due to fiictional resistance. 

Siodda states that the loss of energy in steam turbine noszles is approximately 15%. The actual 
or expected exit velocity (wO using this figure will be: 

VJi -^24 V (ti -if) (1. - 0.15) « 0.9221U, (lO) 

vi - 0.922 X 2824 « 2604 ft. per sec. 
0.264 X 5.19 



Am 



A^' 



1485 
0.264 X 22.2 
2604 



0.00092 sq. ft. corresponding to 0.41 inch diameter. 



« 0.00225 sq. ft. corresponding to 0.64 inch diameter. 



Design of Blades. The peripheral velocity (c) of the wheel is limited to the figures given by 
Table 2, (ot safety in operation. The centrifugal force developed by higher velocities produces 
stresses in the wheel that cannot be well taken care of and still provide for a fair factor of safety. 
In practice a blade giving a complete reversal of the jet, as shown by Fig. 4, cannot be used as 
the steam leaving the blade must clear the wheel. The steam must therefore for practical reasons 
enter and leave the ^eel at an angle. 

In the De Laval single pressure stage turbine the nozzle makes an angle a « 20^ with the 
s — X axis of the wheel, as indicated by Fig. 6. 



POWKR PLANTS AND REFRIGERATION 



3i and angle of exit h of the bladea (Fig. 9) a. 
Theae angles vaiy in magoitude from 28 to 3S 




Flo. 8. Detaii^ or Biades. 




Fio. 0. Blasb 8>cnom.* 

X— X axis (A wheel. The backs <rf the blades are made parallel with the ang^ lines. T 
thicknesB I at th« blade edge ie made approximately 0.02" to 0.04". 

The front irf the blade is a circular arc the radius of which is found by dropping a pep- 
dicular from the edge of the blade to the center line. The foUowing blade or buoket di' 
Tal^ 3, may be used for the use of machine indicated. 



Ill 



1 

11 I!! 

111111! 



III 



306 



POWER PLANTS AND REFRIGERATION 



TABLES 



Hp. Rating 


Height 

of Bucket 

inQew 


Width 
of 

Btiekst 


10 


0.60" 
1.10" 
1.40" 


0.40" 


100 


0.40" 


600 


0.60" 







Velocity Diagram. Referring to Fig. 12, the nozzle directs the steam on the blades with 
an absolute velocity <^ vh. The wheel is moving with an absolute peripheral velocity oi c. The 
steam enters the blades with a velocity relative to the blade equal to the resultant of tvi and c or trs. 

The angle Pi formed by t^t and the x — x axis is the correct entrance angle of blade to avoid 
shock. If the friotional loss in the blades be neglected, then the exit velocity tct relative to the 
blades is equal to the relative inlet velocity wt. 

In design the loss of energy in the blades for a single-pressure and single-velocity stage turbine 
may be assumed as being equal to approximately ^t ^ 24.7 % (Stodola). 



W Vh* 
The kinetic energy of the steam entering the blades is 



2g 



The kinetic energy of the steam leaving the blades is 



2g 



(1. — i^j)— r — from which 



'^ 



tr, = tTi V (L - !^i) 



(11) 



The loss of velocity in the blades may also be approximated by means <^ the curve, Fig. 13. 
The curves shown by Figs. 13 and 14 were taken from " Notes on the Curtis Turbine," by Lieut 
0. L. Cox, U, S. N, 

The component of uti in the direction of rotation is t^'i = u^ cos a (a ^ noszle angle). The 
velocity of the jet rdative to the blade in the direction <^ rotation is 

w\ ^ COT Wi cos a -^ c 

The impulse on the blade in the direction of rotation or blade motion is: 

W W 

Fp = — (u/i — c) = — (u^i cos a — c) lb. 

9 9 

The reaction produced by the jet on the blade in the direction oX. motion b 

W W 

Fr ^ — (w'i + c) = — (Wi coe 4> + c) Vb. 
9 9 

The total force produced by the jet acting on the blades is theref(H« 

W W 

F = Fp '\- Fr '^ — (w'l + u/i) = — (u^ cos a + W4 cos 0) 

9 9 

The en^iyy absorbed by the wheel per a. o/«(eam (TT = 1) or useful work is 

c c 
• ^c = — {w'l -\-wU) = — (u^ cos a -h ti74 cos ) ft.-lb 

9 y 



(12) 



The heat equivalent of the useful work is 

zizT oXm, per lb. <rf steam ]' 



(13) 



STEAM TURBINES 



307 



The efficiency of the nozzle and wheel is 

energy abeorbed by wheel per lb. of steam 



E 



Available energy oi the steam per lb. 
Fc 2c (u/i+ w\) 



v>.*/20 



w^ 



(14) 



There is a further loss ^t due to windage, leakage of steam past the buckets and mechanical 
friction. 




Nozzle 



p « 81.2* Theoretical exit ^el. from nozile w^ - 2822 ft, per se^. 

H'^w, Cosa^244& 



^0.5576 




CondWons 
Inftial pr, 125. i^ gage 
T^mlnalpr, /' " 

Infi/al quality X'* f.O 
Nozzie loss 16% 
Blade " 24.7% 
Peripheral Velocity 1120 ft. per sec 
Energy absorbed by wheel per lb. steam per sec. * 

--^ (2445 -^87)'^ 86230 ft. IJ3S. 

Efficiency of Rozzle and Blade, 
E^JBiO^ ^0.89 01-69% 

2824 * 

2g 



Cos fir Cos ^,".8295 



w, Cosfi^G^SI 



Fio. 12. YBLOcrrr Diaqram for Sinolb Staob Impttdsb-Turbinb. 

This is quite variable and depends upon the size, design, etc., of the machine under considera- 
tion. The oombination is stated by various authorities as 38 to 80% of the useful work as found 
by equations (12) and (13), the lower figure used for non-condensing and the higher for condensing 
machines. The brake horsepower developed by the machine per lb. of steam per sec. is 

-p-.'^a^ <«. 



The estimated steam consumption per brake horsepower-hour or water rate of the turbine is 

3600 1980000 



WR. 



d.hp. per lb. per sec. Fc (1.— ^i) 



; lb. per hour 



• 



(16) 























































o 
c^ 












^ 










































^ 












> 


V 








































Ov 














N 


V 






















































\ 






































Oe% 


















\ 






















































^ 


\ 


































^ 




















\ 


V 




















































\ 
































^ 
























\ 




















































> 


\ 




























*o 


























\ 


L 




















































\ 
























































\ 
























^ 






























\ 


V 






















ftN 
































\ 








c2 














*^ 


































\ 






«0 














«M 


































\ 






o 
















































\ 






















































> 


V 


1 




















































\ 


mj— 






















































V 














^ 






















.•S" 


















\ 




































1 


















^ 


V 












^^ 




















.4 


C:- 




















\ 












OQ 






















5> 






















V 


























^k 


t 


to 

-«o 

o 

-J 


iz 






















\ 










K. 
















s 

• 


i 






















^ 


V 
























is 
























\ 








to 


















1 


1 


1 
























> 


L 




















































\ 






Ui 






































































































V 




^J 


















































\ 




V 


















































\ 




oo 




















































[ 




















































\ 






















































\ 




















































\ 






















































1 


o 




















































1 



§ 

o 
** 

CO 

V. 

-^ 

O 

o 

CO 



<0 

i 



I 



c:> 
c:> 



CM 



CM 






c:> 
c:> 



§ 



puooffs ^ i^J 'Apoioj{ (// 9^07 



Ov 




§ 




Ok 










i 




\ 




Los 


i/fl Ks 


ocity: 


Feet p 


?r Sea 


nd 










Kn 




\ 


^ 




























\ 




























\ 






















Cx 






\ 


\ 










1 II 


1 








Qk 








\ 












4k 










^., 


5s 






\ 










Co ^ 

5 r^ 










Co 

^ 


1 






\ 


\ 


























? 


\ 




• 




Q- 










3 








\ 


k 
























Co 




\ 
















S ^ 

Q 

c 

! K. 

i 

• 








1 




\ 


\ 














1 












\ 


I. 












•1^ 


Q. 














\ 


























\ 


\ 










<& 


















\ 


k 








Vj 




















\ 




























\ 


\ 


























\ 


v 




% 


» 






















\ 


, 



310 POWER PLANTS AND REFRIGERATION 

Example. Required the blade angles and estimated steam consumption for a single stage impolae 
turbine to develop 100 brake horsepower and to be operated under the following conditions: pi ■■ 140 
lb. abs. pt « 15.7 lb. abs. steam initially dry and saturated, peripheral velocity of wheel c «> 1120 fU 
per sec Assumed losses \pi » 15%. \pt >- 24.7%. ^i » 39%. 

Nossle angle a * 20**. 

The theoretical nossle eodt velocity Wg « 2824, as given in the previous example for the 
conditions of pressure. 



wi 



- Vl. - .15 X 2824 - 2604 



Referring to Fig. 12, the absolute nossle exit velocity wi is laid off to scale on the ndsale angle 
line and comluned with the peripheral velocity c » 1120, as shown. 

The component of in in the direction of blade motion is lo'i - 2604 X cos 20® - 2604 X 0.939 
» 2445 ft per sec w'l — c «> 2445 — 1120 ■» 1325. The vertical or axial component of wi is n^ « 
101 sin a - 2602 X 0.342 - 890. 

cot/8i - ^^"^ - ^ - 1.4887 oi-^i - 33* 65' 

w'l — c 1325 

Entrance velocity to blade relative to the blade is: wt — — — ^ ,,,^^ * 1596 ft. per 

cos^i 0.8295 

The relative velocity at exit from Uade is: 



Wt - 1596 X V 1. -0.247 - 1388 ft. per sec 
c + v'4- vi cos/St - 1388 X 0.8295 - 1151. 

.*. w'i - 1151 - 1120 - ai. 

The energy absorbed by the wheel per lb. of steam from equation (12) ii 

Fe - i^ (2445 + 31) - 86230 ft-lb. 

The efficiency of the nozzles and blades is 

2824* 

The estimated steam consumption from equation 16 is 

1980000 
'^•*- ■ 86230 X (1.-0.39) '^^^-^ ^rake ho»>pow»l>oar. 

The height of the blades using the exit diameter of nossle 0.64'', calculated in the previous problem, 
may be made l.O''. The width of the Uades may be made H''* The pitch of blades may be made ap* 
proximately ^ of the width or ^'' on the pitch line of wheel. The speed of wheel may be taken from 
Table 2 or approximately 13000 r.p.m. The pitch diameter of wheel is therefore equal to 

1120 X 60 
D - ^ - 1.645 ft or 19H". 

T X 13000 

IMPULSE TURBINE WITH VELOCITY STAGES 

The single stage wheel, owing to the stresses produced in the blades due to the hic^ roiadye 
speed necessary, limits their height and consequently the area for the passage of steam. The 
maximum capacity <^ a single wheel due to this limitation is approximately 700 hp. 

If the capacity of the turbine is to be increased it is therefore necessary to increase tlie num- 
ber of velocity stages. By means of multi-stages the speed may be sufficiently reduced to pomit 
the use of the longer buckets required to pass the larger volume of steam necessary to devetop 
the greater hcHisepower. This may be accomplished in several ways. 



STEAM TURBINES 311 

(a) Hie expansion may all take place in a single set of noEiIes (single preaeure stage), with 
wrend sets of wheels with stationary guide vanes between the wheels, aa shown in Fig. IS. The 
Atberger turbine (Fig. 16) and TViry turbine (Fig. 17) belong to this class. 

(b) Tbe preesure may be divided among several seta of nosclee, one aet for ecu^ etage, each 
prtmim stage having two or mcwe velocity stages as in 

the CwtU turbine (Fig. 18). 

(c) Multi-pressure stage with one velocity stage 
(one wheel) for each i^eeaure stage as in the Bateau 
Untmie (Fig. 19). 

Impulse tuibines with a plurality of pressure stages 
(b cr c) have received the name BoM-ceUvlar. 

Velocity compounding or stagjng permits of a 
leductMxt in wheel speed bo that electric generators 
may be direct connected to the shaft, thus avoiding the 
use of reduction gearing necessary when single stf^ ma- 
Gtmes are used for this purpose. 

Sin^ ^eianre — Tlire»-VeIocit; Stage Impulse 
Tnrbme. In this tyi>e the entire expansion takes place 
b coe aet ot nociles as in the single stage De Loral tur- 
bine. The steam leaving tbe blades is re-nlirected upon 
the f oOowing set of blades by means of guide vanes, as 
shown by Fig. 2a 

Energy Loss in Blades and Guide Vane Passages. 
A losB oS kinetic energy occurs both in the blades and 
guide vanes due to the frictionat resistance offered to the 
pamage (rf the steam. 

Let uv = velocity at inlet to blades relative to the 
blade or the absolute velocity at inlet 
to guide vanes. 
u^ E- velocity at outlet of blades relative to 
the bUde or absolute velocity of out- 
let from guide vanes. 
4>i ~ fractional part of energy lost by friction 
in blades and guides. 

Pta. IS. 
whed or guides. Vblocit* Staom. 

■n^n ^ X (1.0 - W - ^ 

.*. K. - Wi V 1.0 - ^, (17) 

An energy kiss of 10% (^i — 0.10) in both blades and guides is frequently assumed in tenta- 
tive dedgns lor two or more velocity stages. This loss is a function of the velocity and is there- 
fvenot aconstsnt. Several formube have been proposed by which this loss may be estimated, 
ooe of which is given by Figs. 13 and 14, and may be used ic 



Conotruct the velocity diagrams for a mn^a-preesure tbree-velodty stage turbine to 
beoperated under thefolIowingcoDditiDna: Initial praoure, ISO lb. gage (pi - 164.7) ateam diy saturat- 
td, tcnmnal prenurB 28" vacuum (pi - 1.0). Speed of wheels. 3500 r.p.m. Pitch line diameter of 
•U whOeb 3' - 0". Peripheral velocity e - 550 ft. per sec. Assume ^i - 15% for noule and ^i 
- 10% for each wheel and each set of guide vanea. From the lioUier chart i, — ti ■• 320 B.tu., 



312 



POWER PLANTS AND REFRIGERATION 



theoretical noiile eodt velocity to« - 4040 ft. per see., m - V 1.0 — 0.15 X 4iHP "" 3717 ft. 
estiinated actual exit velocity. 

Referring to Fig. 20 the velocity diagrams shown and the figures given in moat caso a 
tained by simply scaling the diagrams and are therefore subject to slight corrections which 



ob- 
be 




Fio. 16. Albsbobb Two-Stags Turbine. 

obviated by solving the triangles as was done in the previous problem. The exit angle of the guide 
in each case was made equal to the blade angle of the preceding wheel. 

The total energy absorbed by the three wheels per lb. of steam supplied is given on the figure and is 
174,320 ft-lb. 

Allowing a loss due to windage, leakage and mechanical friction of ^t ■» ^%t the Bteam per brake 
horse-power hour will be 

1,980,000 



W.R. 



0.80X174^20 



14.2 lb. 



Few Pressure Stages with Several Velocity Stages for Each Pressure Stage. The CwUs 
turbine in the larger sizes is built with four preasure stages, each pressure stage having two 





Fio. 17. Abbakgbmsnt of Buckstb and Rbvbbsino Ohambbr Tbrbt Tuasm. 



stages. The smaller sizes are constructed with only two pressure stages, each pressure stage 
having two velocity stages. One of these machines is shown in section by Figs. 21 and 22. 

If the heat drops (ii — t't) for each pressure stage are made equal, then the velocities of the 
steam entering the first wheel of each stage will be the same, and if the blade angles of the wheds 
f (N- each pressure stage are alike the energy absorbed by the wheels f cht each pressure stage will 
be equal. 

The action of the steam is conveniently studied in connection with a MoUier diagram. Re- 
ferring to Fig. 23, let the initial state* of the steam entering the first stage pi be indicated by 
point A on the diagram. Under ideal conditions, if the kinetic energy of the steam was all ab- 
sorbed by the turbine and the terminal pressure was pi then point D would represent the final 
condition. The length A-D would then represent the energy absorbed and w(»k done. If A-D 
be divided into several equal parts as ilB, BC and CD then the heat drops are equal and the 
initial pressures for each stage as pt and ps are known. 

In the actual turbine, however, the wheel or wheels oi each stage do not extract all of the 
available energy, as represented by the lengths AB, BC and CD, as there remains in each case 

the absolute velocity oi exit from the blades which carries away -- ft.-Ib. of watk per lb. o£ steam 



STEAM TURBINES 



313 



aqiplied, and in wlditHXL tha« is the loas due to frioticHi. Let j4£' represent the heat avEiilable for 
the fini premiTB etage. 

If the length of the segment AF repraeente the heat equivalent of the energy abeorbed by the 
fint Btage as determined bom the velocity '^i|^e^w^nH then the length FB' represents the heat 
equivaleot of the loos for the Gist {wessure stage. 




lliis heat, exoqtt for a sm&ll fraction that is radiated, is expended in raising the qtmlity of the 
•t«ain (et npeilwating it). Hie quality (or supo'heftt) at the end oi the first stage will therefore 
be fooDd by drawing the hwisiHital line FE to the intersection with the initial preosure line pt for 
the Meond stage. Tliis point (E) represents the initial condition as to pressure pi and quality 
of the steam lot the aqmnsion nosslea or TUtes for the second stage. 



314 



POWER PLANTS AND REFRIGERATION 



'The heat equivalent of the work absorbed* by the second-stage wheels is now laid off <m the 
veitioal tta EH; pass horizontally to the right to the intersection with the initial pressure line p» 
for the third stage, the condition being represented by the point /. Owing to the fact that the 



Theor. exit yel.nozz/e=4040. 



Theor, Jet velocity m224^ 326 - 4043 

Nozilelo$$^l5f 

Actual Jet vel.»4043^l.-.l5^37l\ 

Aiiumed lose In each row et guide 
venee end bledee »IOfi 




exit velocity for guide vanes and blades 
ss Inlet velocity a^I.-JO 
JB " " X 0.948 



Energy absorbed by wheels per lb, steam 
fat wheel ciw\-w') = 98054 ft. lbs. 

2fld wheel c(w'^w') ■= 55917 ft. lbs. 

Bra wheel c jwl'-w'^ = 20349 ft. lbs. 
S Totals 174320 ft. Iba, 

Efficiency of nozzle, Guldea ^ blades, 

^^, 174320 ^ ^Qg (nearly) 
4040} ^ ^' 

29 



Single Pressure StageVlhree Velocity Stages 



Fig. 20. YELOcnT Diaqrams fob a Sinolb-Pbesbure. THRES-vBLOcrrr Stage Tttbbenb. 



pressure lines are not parallel but slightly diverging, equal heat drops for each stage will not 
result by using the pressures for the several stages as determined by dividing the total heat dn^ 
(ii — is) I line ADf into equal parts. 

Reheat Factor. The heat drop per stage, however, which will accomplish this result ap- 



STEAM TURBINES 



Fra. 22. Detail of Ciibtu TwoStaoe Tubbmb. 



^ 



316 



POWER PLANTS AND REFRIGERATION 



proziinately is determined by multiplying the theoretical heat drop (ii — it "■ H — «•*«§— ig^ 
etc.) per stagei as determined biy dividing the line AD into a number of equal parts oofreqxMiding 




Entropy 
Fig. 23. Heat Drops Oxtbtis TuRBDnB. 

to the number of stages to be employed by a factor 1 + li^* The value of iiC is found by the 
empirical formula following (F. E, CardvUo, Tram. A. S. M. E., 1911): 



K - 0.00056 (' — -^ A F (1. - J&) 



STEAM TURBINES 



317 



AH 

n 
B 



total avaOable heat drop (ii — ii) 

number of stages. 

probable internal efficiency of each stage (about 60 to 70%). 

Let it be required to lay out the blading for a Curtis type turbine having three pressure 
pressure stage to have two velocity stages. Initial pressure pi » 105 lb. absolute, terminal 






Theor» exit nozzle velocitf=224 VQ\~Qa 

==224 >/IIT^2384 

c=400 A ctual expected exit velocity=2384 \J L-'J0'^22S2 

w;^, CCS. 20»2l2e 




Energy absorbed 1st wheel 

^jOO (2/26-^ I28T) ^41828. ft. lbs 

9 . 

Mosses 

lOf Noii/t 
** Bl»d§9 

if euid§ 



/5=ft.2¥--/ 



EfncfsncY 
£^e2m^Qj05 




9 




Energy absorbed 2nd wheel 

^400 0246-^402) =20486. ft. lbs. 
9 

Total ^62313 ft. Ibsu 
62313 



PrP,^33*-2r 



778 



mZO B.tu* 



WjmmW^\/LO—O.I 



per see. 



Pia. 24. VBLocrrr Diagbam fob Each Pbbbsubb dTAon of a Oubtis Ttpb Tubbimb. 
P4 - 1 lb. Total heat drop %x - ii « 326 B.tu. Peripheral vdodty of wlieela c - 400 ft 

K - 0.00056 (^-^) X 326 X (1 - 0.70) - 0.04 nearly. ^ 
Reheat factor 1 + it - 1.04. The theoretical heat drop per stage is — » 109 B.tu. The 



318 POWER PLANTS AND REFRIGERATION 

heat drop per sta^e which must be used to obtain an approximate equal division of work in each stacs 
will be 

109 X 1.04 - 113.4 B.tu. 



Theoretical nosile exit vel ocity fo r each stage nossle Wt » 224 V 113 » 2382 ft. per sec Exi>eoted 
velocity wi » 2382 X V 1. — 0.10 » 2264 ft. per sec. The expected heat drop per stage is, 113.4 X 
0.70 = 80 B.t.u. (nearly). This is the energy absorbed by the wheels of each stage. The loss of 
energy in each guide vane and each blade is assumed as, V't = 10%. 

The velocity diagram for each stage is shown by Fig. 24. The exit angle fi* of the guide vazie in 
the construction shown is equal to the blade angle fii ■■ fit of the preceding wheeL The p rcwn ires 
and qualities for the 2nd and 3rd stage as determined by means of the MoUier diagram as previouflly 
explained (Fig. 23). are ps « 37. xt « 94.1%, pi » 6.8. xt - 89.6%, pa « 1. 

The nosile calculations are similar to the example given for the single-stage turbine. 

The total energy absorbed by the wheels per lb. of steam, assuming a blade and guide vane loss 
i^t " 10%. is 3 X 62313 ft.-lb. 

Assuming a loss of 20% for windage and mechanical friction, the estimated water rate is 

1.980,000 

1FJJ. « «■ 13.3 lb. per hour. 

0.80X3X62,313 *^ 

IMPULSE-REACTION TURBINES 

A pure reaction turbine is one in which the energy is derived from the reacticm due to the 
expansion of steam in nozzles or blades attached to the wheel or rotor. 

No pure reaction turbine is now on the market. A combination of the impulse and re- 
action principles, however, is employed by one of the most important types of turbine thus far 
developed. 

Fig. 25 shows the sectional elevation of a Weatinghouse-Paraons ungle flow tiurbine em- 
plosring a combination of the impulse and reaction principles. In this type of machine no nozzles 
are employed, the expansion of the steam taking place in each set of stationary guide vanes and 
yioving blade passages, steam being admitted around the entire periphery of the rotor. 

The steam is admitted to the tiu'bine through a poppet valve actuated by a governor. 

Owing to the large number of rows of blades and guide vanes employed the heat drop per 
row is comparatively small with the result that the steam velocity is also low, the steam velocityi 
varying from 150 to 600 ft. per sec. 

The complete expansion of the steam is carried out in the annular ocHnpartment, formed 
by the blades and guide vanes, which resembles in effect a divergent nozzle. 

The impulse-reaction turbine allows the lowest velocities of rotor acocHnpanied by hi^ 
economy that have been attained. 

The absolute velocity of the steam as it leaves the guide vanes varies progressively in passing 
through the turbine. The exit angles are made the same for both guide vanes and moving blades, 
ordinarily between 20° and 30**. 

When the absolute velocity of the steam leaving the guide vanes would rise above 2 to 3H 
times the peripheral velocity c, the diameter of the rotor is increased. The following table gives 
some particulars as to speeds and number of rows of blades used in the Pantona turbine. (E, M. 
Speakman.) 

The action of the steam in a turbine of this type is shown by Figs. 26 and 27. A common 
ratio of peripheral speed to the absolute velocity of the steam leaving the guide vane is approxi- 

mately — » 0.60. Having selected the peripheral speed from Table 4 for the first stage it is 

necessary to select the exit angle a (20^ to 30°) and construct or compute the entrance angjb fl 
for the blades. The heat drop per row and stage may then be calculated. 

A typical arrangement for the rotor in this type of machine is shown by Fig. 26, the rotor 
being divided into three cylinders. 



STEAM TURBINES 



POWER PLANTS AND REFRIGERATION 
TABLE 4 

VARIOUS VANE 3FGEDS 





P»P»».V^8P— 


"'bS:-' 


STffl: 


Kontal Output ol Turbioe 


nnt 


LMt 






Eipu^u 






KWOkw 


SE 


810 


70 


TW 


SSOOkw 


ts 




TK 


UOO 


ESOOkw 


ts 




M 


lUO 




UG 


ICO 


re 


IM* 


1000 kw 




»iO 


so 


ISOO 


TMkw 


12B 


260 


77 


200D 


SOOkw 


120 


28e 


SO 




EMhv 


00 


110 


n 


«000 


IGk* 


00 


too 


48 


4000 



A group of blades having the BOme hei^te is termed a band, so that the typical 
is thiee cylinders with three barrels to each cyUnder. The increasiiig height (A bUd 



Pro. 2a. WEBTIHaHODBE-PAIMONS TUBBINB — MOIAI-PrXHSOBX — MlTUn-TBLOCITI STAOK. 

to the increase in the volumo of steam as the pressure decreaaes. The diameterB of the cylinden 
increase in order, to prevent excesHive blade height which would otherwise be aeocaoiuy to provide 
sufficient area hi pass the large volume of steam at the lower preasurea. 

Ezunpls. Let it be required to determine the number of rom of blades (or aD imfiiilsn rnnr linn 
turbine lo be operated under the following conditioas: pi ~ 165 lb. absoluts, pi • I lb. abaolute, total 
heat drop ii — it " 326 B.t.u. Rotor to be composed of three cylinden, speed from Table 4. lit 
cylinder (small) e — 136 ft per sec.; 2nd cylinder (intermediate) c — 220 ft per see.; 3rd cyliader (laixe) 
- 330 ft. per seo. Exit angle of guide vanes and blades a — 22H*- Ratio of pvipheral qieed to 

135 
For the fint cylinder tui - r—- - 22S (t. pet aeo. 



As erpansion also takes place in the blades, aa well as in the guide vanes, the exit velocity 
from the blades iri, relative to the blades, may be made equal to the absolute exit velocity uh from 



PiQ. 27. AcnoN OP THa Stkui m to* Pabbonh Tcodin 



e — BiMde Tilocltr at mean diamcrcer. 



oAniidSCatNoliita 







Fid. 28, TBLocnr Diaohah for luFOTOK-RiAonoK TuBsnn.' 



322 POWER PLANTS AND REFRIGERATION 

the guide vanes, or wi =■ toi. Constructing the velodty diagram, Fig. 28, we find ui'i » 20S and 
V!', = 70. 

The energy absorbed by one row of blades or one stage is: — — (205 + 70) = ll.M rt.4bi. 

The heat drop per stage is 11.54/778 or I.4SB.t.u. Owing to the loss by friction and leakagB the 
actual heat drop will be greater per stage than the theoretical. 

If this loss is assumed as 25% then the actual heat drop per stage for the first cylinder ia 

'- = 2.00 B.t.u. (nearly). If all the stages were on the small cylinder of the roUx there 

would be required 326/2 ■= 163 stages. If the rotor be divided into three cylinders and ^proxi- 
mately equal work is to be performed by each cylinder, the number of stages for the first cylinder ia 
163/3 ' 54 + ■ The number of stages for the intermediate and large cylinder may be determined 
in a similar manner. 

In Westittgktruae turbines for small powers the moving wheel, ix rottv, is a sted '•"*'"& 
with blades inserted in a groove cut in the periphery (A the wheel, and held in place by double 
or triple rivets in each blade, according to aixe and speed. The cast-iron cylinder is ^lit hocixoD- 
tally, a section through the bottom half of a turbine cylinder being shown diagrammatically in 
Fig. 29. 

Steam entera the turbine at A, throu^ a valve of the double-seated poppet type; passing (be 
first stage nozzles B, it impinges on the moving blades, giving up part of ita energy, thoice paooes 



Fio, 29. WnrrnraBoms — Two-Staqi Iifpriaa Ttmsim. 

to the first reversing chamber C, where its direction is changed and it is made to pass again throuj^ 
the moving bUdes, giving up the balance of the energy imparted by the first stage noules. The 
st«ain then passes through the second stage nozzles D and the second stage reversing chamber S, 
where the cycle above described b repeated, finally leaving the turbine through the exhaust pipe F. 

Turbines of 50 hp. or leas have but one set of nozzles and one reversing chamber; all oUieia 
have two stages. The nozzle blocks and rcvetsing chambers are made of bronze and finished by 
hand to insure the minimum of friction losses due to steam passing over these surfaces. 

Two general types of governor are employed. For turbines driving centrifugal pumpa, fans 
and the like, where exceedingly sensitive regulation is not required, a shaft governor is used. 
On the larger machines for electric drive, a govemiv of the flyball type, driven from the turbine 
shaft by bevel gears, is used. With all turbines an automatic safety stop is provided, which shuts 
off steam in case of overspeed. 



STEAM TURBINES 323 

The f<41owuig clanee irf tmbinea are maDuf&ctured by the De LmxU Steam TmUne Co.: 
CUn " A " Sinj^e-«Uge Impube Turbine, rangiDg in capacity from 7 to 600 honepower, 
of R smgle get of noules discharginK steam upon buckets of a single high-speed wheel, 



the power being transmittod to the driven machine by means of helical redueticm gean 
(Fig. 10). 

Claas " B " Singl»«tage Impulse Tuibioe, ranging in capacity from 6 to ISO bonepower, is 
deaigned tot driving extra hi^t^peed machinery without the intennediation of gears. 

Class " C " Velocity-stage Impulse Turbine, raiding in capacity from 1 to 600 horsepower, 
oontaine a nn^ set (A nozilee ia which the steam is expanded from the initial to the terminal 



324 POWER PLANTS AND REFRIGERATION 

pressures. From ttiese noiiles the steam ia discharged against a row <rf moving budcets and is then 
redirected by stationary guide vanee upon a seccmd set ctf moring buckets. In some caaee a aeocHtd 
set of stationary guide vanes and a third set ot moving bucketa are employed. Tbe Class " C " 
Turbine is peculiarly fitted to operate with high-pressure steam exhaustins to atmosfriiete or 



PlO. 31. DniENBION Drawinq Of Stctrtevant Turbink. 
TABLE 5 

DIMENSIONS OP STURTEVANT STEAU TURBINES 



.SE, 



ISM 
26 M 



'iii 



2HTU 






3 Phut. Not ont 8600 Vdl& 







28 ft. 


















































































































OiB. 


lilt. 


lin. 




4tT,IXII> 





SG 


CrdMlorSPluH 


Not ovw 6600 Vdt* 






300 






Sft, 41n. 




9iD 








































































































760 




18 a. 8 In. 


1 ft. 


6ln. 





ffTEAM TURBINES 325 

a^inst back-presBure ix with low-pressure steam exhaustiog to condenser, or for use as a inixed- 
flow turbine. It is deeigned for direct coniiecticm to centrifug&l pumpa aiid blowen, sm&ll alter- 
nsting or tiireet-curreat eeDer&tors, centrifugal air compresson and other moderate ih' high-speed 
macfainery. It is also used with gears tar belt or rope drives or for driving shafting or slow- 
^Med machine^. 



ClBaa " D " Pressure-stage Impulse or Multicellular Turbines {Fig. 30), ran^ng in capacity 
front GO to lS,O00hp., consist of a aeries of Bingje-stage wlieels, each enclosed in a separate cell or 
compartment and all mounted upon a common shaft. This turbine is directly connected without 



POWER PLANTS AND REFRIGERATION 



STEAM TURBINES 





J' 


ll 


— '-3 = 

iiiii 




1 






' 


" 


m 




• 






' 


4 
Sssa 




■ 


ii::: 




1 


!>.»•• 




- 


:feil 




X 


:k«al 


3 




».f„S 


„ 


fttf? 




= 


!??:= 




. 


IS: 




. 


5s2aSR 




» 






» 


^TTT? 




. 


S,««f« 




' 


JTTT? 








££JSMt4 
Hill 



328 



POWER PLANTS AND REFMGE5ATI0N 




S 



U9 

CO 






00 



I 



M 



n 



a 



•t 



v4 vrt vN^4e« A 



e«c4eo< 



1:2999:3199 




i I I I I I I I 
<« iO «e t« t« «D e» o 

rr?T I I 1 1 1 

C9comc4eoeO'«<«i 

^0000000 

^1 I I I I I I I 
I I I I I I I I 

as 

I I I I I I I I 
McocoeocoeO'^iC 

J I I I I I I i 

i 

c4 COCO coco cocoas 

-a — 

^ ^ pbl (1« pbl pbl CL4 (X4 

COCO CO coco coco CO 

SSSSSSSS 

11 

888SS8SS 

CO 00 GO C9 coo CO CO 
■^ '.^ • . . . . 

t^ »-< el « « «o r* 



STEAM TURBINES 



329 




g 



o 
& 



CO 



I 



M 



W 



O 



H 



O 



A 



I 



-fl-fl j-g a-fl 

•0>Q ■•O fj>^ 



5 ••^»*\ 'p^ 

NMto '0000 'le 



C0 + 






3? 



:ogMo 

■J ^ « 

CO ^lO 



+ 



#:st^ 



8 to 
CO 



J 



•00 






i- 



J I I I I I I i 

«DOOOOMM>OeO 






"•f 00 



Mil 
ooee»- 



5 :s; ;s;xj^;s::« 

leo^ooioooeoee 

I I I I I I I I 
Mcoee^^^iof 



O0k«HC0Q0iH«-4O 
J I I I I I I I 

eoee>o»e^^^«D 






m * * 

leoQieieeeo 

fHSSfc-MQQQ 



i 



i 



1 



1 



330 



POWER PLANTS AND REFRIGERATION 



gears to high-speed machinery, such as alternators, compressors, high-speed centrifugal pamps 
and with gears to slow-speed machinery, such as direct-current generators, large o^itrifugal fans 
and pumps, heavy machinery and for rope or belt drive. Special small-wheel turbines are buiH 
for back-pressure service. 

LOW-PRESSURE TURBINES 

A low-pressure turbine is generally understood to mean a turbine designed to operate on 
exhaust steam at approximately atmospheric pressure. 

The most general application of the low-pressure turbine is in connection with non-condenaui^ 
engines, the exhaust from which being piped to a turbine, the latter opiating condensing. 
Fig. 38. 

The available energy in one pound of dry and saturated steam between the limits of 140 lb. 
per sq. in. absolute initial pressure and 16 lb. per sq. in. absolute terminal pressure corresponding 

TABLE 10 

DIMENSIONS OF CURTIS STEAM TURBINES 



Rating kw. 


Voltage 


R.P.M. 


Length 


Weight. 
Pounds 


16 


86 
125-260 
125-250 
125-260 
125-260 


4000 
8600 
2400 
2000 
1500 


6'-4J" 

e'-o" 

18'-0" 
16'-0" 
17'-0" 


1850 


25 


3600 


75 


12000 


150 


25600 


300 


80000 







to the range that may be used in a non-condensing engine, t'lM — iw "" 1194 ~ 1035 «» 159 B.t.a. 
The available energy in one pound of steam leaving the engine at 16 lb. pressure and a 7&" vacuum 
or 1 lb. absolute corresponding to the pressure range, as used in low-pressure turbine practice is: 
tie - ii = 1035 - 877 = 158 B.t.u. 

It is thus apparent that when condensing water is available the capacity of an existing non- 
condensing engine plant may be practically doubled by the addition of a low-pressure turbine 
unit, and the necessary auxiliaries, requiring no increase to the boiler plant or increase in the fuel 
consumption. 

When condensing water may be obtained at the cost of pumping ht>m a nearby supply this 
combination has proven in many cases a very profitable investment. 

The diagram. Fig. 37, shows the performance of a Rice-Sargent engine coupled to a 1200-kw. 
direct-current 250-volt gaierator. The upper ciure shows the relation of load to water rate non- 
condensing; the intermediate curve the same relation with 28" vacuum; and the lowest curve the 
steam consumption at various loads of a combination of this engine with a low-pressure Curtis 
turbine. The ciures show a net water rate of 16 lb. for the combination as against 22.6 lb. for the 
engine running condensing alone at 1200 kw. It also shows an increase of peak ci4)acity of more 
than 1000 kw. due to the turbine, and finally, that, using the same amount of steam as is required 
to run the engine condensing at 1200 kw., an output of 1710 kw., that is, 42.5 per cent more, is 
secured. Looked at in another way, with a steam flow about 50 per cent greater than that re- 
quired by the engine when running direct to the condenser, the combination of engine and turbine 
will give an increased output of about 100 per cent. 

In non-condensing plants where there is a sufficient supply of exhaust steam at all times the 
straight low-pressure turbine fulfills all requirements. There are many plants, however, where 
the supply of exhaust steam is not constant or where it may be desired to secure more power than 
can be generated by low-pressure steam alone. 

The method employed in this case is to cross-connect the live steam header with the pipe to 
the low-pressure turbine, the connection being equipped with an automatic reducing pressure 
valve. 



STEAM TURBINES 



331 



800 



34 






— 


"^^ 




■^ 


^^ 




^" 


^^ 




™^ 


^"" 










^^ 


■^^ 




■^ 


■^^ 




^" 






■^" 








^"" 


'*"" 


^^ 


^"" 


"^ 


^^ 








































































































2 






















































92 




~ 1 












S 


2 












T«$ta of 22 X 44 A 42 Rhe <f Sargent Engind 

with D.a 250 Von Oenentor 
Compared with Raau/ta obtainable from same 
engine in connection with low preaaure A.C. 
Turbine Unit Pointa marked 0. X M^ 
Corraapond to equel Steem Flowa. 






























.c< 


1^ 


2 


























u 90 




« 


«, 


c^ 


2A 


f0 1 


ro»i 




















































7. 














^ 28 






























A 








































•' 


P 
















^2e 
























. 




*5 


f 






































• 


>3 


z' 


















































^ 24 
























.i 


/ 




















































§ '* 




















c 


^ 


/ 










































































X 


i^ 








































































^ 




























































^*o 


























































































































































«? 18 






















































































































































-^ 




— 


16 
































E^ 


Qt 


0^ 


aa. 


Lei 




•«M 




.....••Ui 


b 


iC' 


Urf 


- 


- 


^ 


— 


•^ 






























■5S 














— 


Zh 


















1* 


1 
































_J 









































1000 1200 1400 1600 1800 2000 

Load in Kw. 

Pig. 37. Performance C^urvbs. 



2200 



2400 



2600 



A.C. 



3 




\\\ Bua-Bars 



Separating 
Receiver 




FlO. 38/ NON-OONDENBINO ENGINE AND CONDENSING LOW-PRESSUEE TURBINB SET. 



332 POWER PLANTS AND REFRIGERATION 

This valve is set to maintain a prcegure on the delivery side sl^tly above atmoephea^ 
It ie replaced in many coses by a special valve operated by the turbine governor, when tlie speed 
falls below a predetermined limit. 

The following classification of low pressure turbine installations is girm by ^nineu Hodg- 
kinton in a bulletin published by the WtsUimhouae Machine Co., " The AppUcation of Low* 
Pressure Turbines," in which a detailed discussion of the various cases enumerated below appeaiv. 

Case A. A low-pressure turbine taking steam from the exhaust of a reciprocating engine, 
thegeneratM^of each being connected to the same bus baiB and no governing device used (Fig. 38). 

Case B. A turbine or a number of turbines and engines connected similariy to Case A. 

Case C. A low-pressure turbine operating in conjunction with one or more engines as in 
Casee A in- B, except that the turbine and engine-driven generators are (A different electrical 



Fio. 30. 

characteristics. A direct-current street railway generating plant, with alternating current dis- 
tributitm to distant substations, is a good example of this case, the turbine and engine-driven 
generators being tied together by rotary convertera or motor generator sets. Another expediait 
by which the use trf a governor could be eliminated is the connection of the turbine-driven alternator 
to bus baiB upon which floats a synchronous motor belted or direct-connected to the reciprocating 



Case D. A low-preaaure turbine operating on the steam exhausted by a number of a 
pumps, or other apparatus, without any relation betwten the electrical output from the turbine 
and the amount of steam available. In such a case a governor controlling the admission valve of 
the turbine is obviously necessary, as is a reLef valve, permitting any excen of low-pressure steaoi 
to pass to the atmosphere (Fig. 39). 

Case E. A low-pressure turbine operating on the exhaust from engines whidi are carrying 
an independent load, as in Case D. The turbine governor, however, controls a valve whicJi con- 
nects the reciprocating engines with the condenser, imposing on them only enough back pressure 
to enable the turbine to carry its load. The enpnes thus have the benefit of some vacuum when- 
ever the load on the turbine is light enough to require less tiian atmospheric inkt pressure. 

Cue F. A low-pressure turbine operating in conjunction with an engine driving a mill or a 
system of shafting, the output of the turbine being used for motors, lights, etc., and any excess of 
current genn^ted over the electrical demand may be returned to the shafting by using a synoehto- 
nous motor, coupled or belted to the line shaft, and thus acting as a balance to proportion the load 
between the two machines so that the best economy may be obtained (Fig. 40). 

Case G. A low-pressure turbine receiving steam from an intermittently (derating engine 



STEAM TURBINES 




Fra. 41. Ra'^u MixEi>-Pmaeu'itE TmiBniB. 



334 POWER PLANTS AND REFRIGERATION 

such as a hoisting engine or a rolling mill drive. If the intervals in the steam supply are not too 
great, a regenerator may be employed, absrabing the excess supply of steam at one time to give it 
up again to the turbine when the latter demands an amount exceeding that passing from the 
engine. 

Case H. Practically all turbines equipped with generators have a valve which will admit 
sufficient live steam to cany the normal load should the low-pressure supply faiL Such an ar- 
rangement does not, however, give high efficiency on high-pressure steam, since its expansive 
energy is wasted in throttling and only a small amount is recovered from the resultant superheat. 
Case H, therefore, provides what is termed a mixed-pressure turbine, which, in addition to the low- 
pressure section, is equipped with elements enabling it to expand steam from boiler pressure to 
that of the condenser. Such a turbine is so constructed that all the available low-pressure steam 
enters it at the proper point. A mixed-pressure tiubine is, therefore, used where it gives better 
overall efficiency, although it has a poorer economy on low-pressure steam alone due to the dead 
load of the idle high-pressure element. The relative proportion <^ the hi^- and low-pressure 
elements will be determined by the amounts of steam of each class to be handled and the con- 
tinuity with which they are supplied. Such a tiu*bine must be equipped with a governor. 

MIXED-PRESSURE TURBINE 

In cases where the amount of power required is in excess of that which may be generated 
by the continuous supply of low-pressure steam there are several forms of machines available, 
designed to operate from two sources of steam. Machines of this tjrpe have been given the name 
''mixed pressure" (Figs. 41 and 42). To obtain the best results it is essential that the machine 
be designed for the conditions under which it is to operate. The following classification of the 
mixed-pressure tiu*bine is given by E, D. Dickson in the " General Electric Review." 

(1) Turbines designed to give the best economy on low-pressure steam and which are equipped 
with a special valve for admitting high-pressure steam to the low-pressure header automatically. 
This machine will not carry any load non-condensing, and will be very inefficient on hi^presBure 
steam. It may be used where the condensing facilities are reliable, and when hig^-pressure may 
be considered an emergency condition. 

(2) Turbines designed to give the best Economy when operated low-pressure, and airanged 
to admit high-pressure steam through separate nozzles. This machine will give fairly good 
efficiency on high-pressure steam, will carry some load ncHi-condensing, and some overload mixed 
pressure. It will carry its full rated load mixed pressure when there is insufficient low-pressure 
steam, or should the vacuum drop below that for which it was designed This class should be 
used where it is intended to operate a large proportion of the time on low-pressure, or in instalh^ 
tions where the boilers will blow when the engine is shut down, or where there is liability of the 
vacuum occasionally dropping off. These machines will continue to use all the low-pressure 
steam available when operating mixed pressure. 

(3) Turbines designed to give good efficiency at high pressure, and also arranged to carry 
load on low-pressure steam. Machines of this class should be used when it is intended to operate 
continuously, or nearly so, on mixed-pressure, and where there is a limited amount of low-pressure 
steam which would otherwise go to waste. In this machine the admission of hi^-pressure steam 
will decrease the quantity of low-pressure steam that will enter. This means that, should the 
machine have to operate mixed pressure on account of low vacuum, the amount of low-pressure 
steam will be automatically reduced and a greater amount of high-pressure steam will be required. 

Such machines are a compromise between a low-pressure and a hig^-pressure turbine. If 
designed to carry full load when operating either way, they cannot be made to give an efficiency 
as high as that obtainable on tiu*bines primarily built for either hi^ or low-pressure operation. 
These machines can be designed to give a good efficiency and carry full load high-pressure, or carry 
about half load low-pressure at fair efficiency. 

In the mixed-pressure turbine, the speed governor will automatically open the low-pressure 



STEAM TURBINES 



335 



▼aive with a decrease in speed, or a falling off in the supply of steam. By a special pressure 
actofited device, the low-fnessure valve may be made to dose and the hi^-pressure valves open 
aut omatiea l ly with decreasing supply of low-pressure steam. 

In order to allow the engine and turbine imit to operate safely under all conditions, the 
Ndmrn^-Brwood Swing Gate and Check Valve has been used extensively. This type of valve acts 
as a safety valve cm any pipe line where it is necessary that the flow through the pipe shall be in 




Fig. 42. WaanNOHOTTss Doublb-Flow Mddbd-Pbbssubb Tubbinb. 



one direction and where disastrous effects would follow if the steam or condensate was allowed to 
flow into the engine or turbine in a direction contrary to that for which it was designed. The 
general arrangement of piping for a mixed turbine and engine plant is shown by Fig. 43. 

When all the steam is used by the operating units, the steam flows from the boiler into the 
engine, then through the turbine and into the condenser. If there is not enough exhaust steam 
ccxning fitnn the engine to operate the turbine unit, live steam may be admitted to it as shown, 
or if too much exhaust steam is flowing toward the tiu*bine, a portion of it may be diverted to 
flow into the feed-water heater, and a proper amount of steam may be used from the engine ex- 
haust to supply the heating system when conditions require. Such a combination of piping as 
duywn makes ioft flexibility of operation under all conditions of service. 

Under all these conditions of operation, the combined swing gate and check valves, when 
placed in their proper position on the pipe, operate normally as an ordinary stop valve; but, in 
addition, they protect the power plant should conditions change without the knowledge of the 
operator. Fen* instance, if the engine is started, without opening the valve on the exhaust, the 
swing gate check valve will open automatically at a predetermined back-pressure on the engine 
and stay open until it is mechanically closed by means of the hand wheel attached to the valve 
stem as in an ordioary gate valve. 

When the exhaust steam from the engine is used for the heating system, this swing gate and 
check valve is used on the atmospheric exhaust pipe as a back-pressure valve. 

Several applications of mixed flow turbines are described and illustrated in the chapter on 
" Exhaust Steam Heating/' Section I. 

Regenerators. In combined engine and low-pressure turbine plants where the conditions of 
operation are such that the engines are intermittently shut down for very short periods, and where 
the load is a widely fluctuating one, as with mine hoists, rolling mills, etc., the installation of a 
regenerator between the engine and turbine unit (case previously menti(med), provides for a con- 



336 



POWER PLANTS AND REFRIGERATION 




STEAM TURBINES 337 

tinuous supply of steam to the turbine. Fig. 44. A steam regenerator is mmply a Tesael con* 
taining a quantity of hot vater arranged to present a large surface to the entering steam. The 
function of the regenexBtor is the same as that of any energy storage medium, namely, to absorb 
tatrfj leceived mCire or lesH intermittently and to give it up steadily. 

Tlie aetiim at the regenerator depends upon the reduction in presBure over the surface of the 
fiqnid below that corresponding to its tempenture which causes a portion of the liquid to evap- 



fOf Oiltrllii 



onte, the beat required for evapcHstion being supplied by the heat liberated by the liquid in 
having its temperature lowered due to the drop in pressure. 

Ttwiwrfe , Aaniming that a dosed vessel contains lOOO lb. water at a temperature of 222.1°, The 
preaaure of saturated steam eoireepondins to thif temperature is 18 lb. per sq. in. absolute. 

If the prewure is reduced to 17 lb. the correaponding temperBture ia 219.4". The heat liberated is 
tOOO (222.4 ~ 219.4) or 3000 B.t.u. The average latent hoat for this range of temperature and pressure 
is 058.3. The amount of water that will be evaporated or steam regenerated is: 3000/968.3 - 3.1 lb. 
For approzimste neulta te find the weight of water neceesary in the regenerator to liberate one pound o( 
■team divide the average latent heat by the temperature drop. 

R^eneratoTS an ordinarily operated between the limite of atmospheric pressure and 4 lb. 
E^ge. 

The constant flow of steam from the i^enerator to the turbine is equal to the average rate of 



338 POWER PLANTS AND REFRIGERATION 

Tlie capacity o! the regenerator to abaorb tlie engme exhiiust detcnuinea the amoUDt ot n 
required. 

Let W = mnTimiim rate (A engine exhaust at peak load, lb. per min. 

V! = mean or average rat« of engine exhaust, lb. per min. 

S = maximum rate passed -to regeneraKir to be ctmdenaed, lb. per min. 
p. W - w 

t, = initial temperature of the water (212° atmoqihertc pressure). 

ti = final temperature of the water (224*, 4 lb. gage pressure). 

fa " average latent heat between the tempetatures Ig and It. 

C •= weight of water to condense 1 lb. at«am. 

= ■ _^ ■ (appnm.). See curvee, Pig. 45. 
D = weight of water to be oiroulated per min. m^""""" demand. 

-SXC-OP-W) 11^ 
E — time allowed to circulate the contento of the regenerator, minutes. 



WHEN WOBKINO BBTWBBN VARIOUS INITIAL AND Fl 

(In the ype of regenerator shown by Fig. 44, the water is drculated and brou^t into oontaot 
with the Bt«am once every 3 seconds. E — 0.05.) 
F = Capacity of r^enerator lb. water. 

— = number of times per min. water must be recirculated. 



F - DE 
Ezunple. The maximum rate of exhaust flow from a number of redprocatiiig mill enEiDes aW^ 
300 lb. per sec. The average rate o( flow is to ' 80 lb. per sec. If the regenerator is to work betweaa 
the limits of atmoepherio pressure and i lb. sage, required the capacity F of regenerator if f « OJOS. 

D - 80 (300 - 80) X _ = 1,063.932 lb. per min. 

F -DE - 1.003,932 X 0.00 - 03,190 lb. 



CHAPTER XII 

PUMPS 

FUNDAMENTAL PMNCIPLES 

Deflnitioiis. StaHc Head ci a fluid at rest is the vertical distance in feet between the point 
mt w^hich the pressure is taken and the surface of the fluid. 

Pressure Head of a fluid in motion is the height in feet of a column of the fluid balanced 
by tbe pressure (biuvting pressure) existing in the pipe at the point where the pressure is taken. 
The pressure head is measured by the height of the fluid column in a straight tube inserted in 
the pipe at right angles to the flow. In dealing with the flow of air this is termed " static head." 
VdocUy Head is the h^, in feet, required to produce the velocity of flow and is measured 
by the difference between the columns measuring the dynamic head and the pressure head. 

Total or Dynamic Head is the sum of the pressure and velocity heads of the fluid in motion 
at the point where the pressures are taken. This head is measured by the height of the fluid 
oohimn in a tube having the end that is inserted in the pipe bent directly against the direction of 
flow. This is termed a ''pitot'' tube. The d3mamic head at any point in the line is the head 
available for overcoming frictional resistance and creating the velocity of flow in the section 
beyond. The ordinary pressure gage on the discharge pipe of a pump measures the pressure 
head only, while the gage (Hi the suction pipe includes in its reading the velocity head as well, so 
that to obtain the suction pressure head the velocity head must be deducted from this reading. 
The total head against which the pump is operating will therefore be the sum of the gage readings 
phis the difference between the velocity heads in the suction and discharge pipes.* The above 
fact should be b<nme in mind particularly when testing pumps. 

If the flow is stopped and the pipe remains full of liquid, the dynamic and static heads be- 
come equal 

^4^!res8iire Equivalents. The various heads measured in feet are transformed to equivalent 
preoBures in lb. per sq. ft. or sq. in. by the following relations: 
h > head in ft. 

d s density of the fluid (wt. per cu. ft.). See Table 1, Chapter 2. z ' 

P » equivalent pressure lb. per sq. ft. 
p » equivalent pressure lb. per sq. in. 
P - M and p - Ad/ 144. 

For water at ordinary temperature (65'' F.) d » 62.345, p « 0.433 A and A « 2.31 p. 
'^ Units of Measurement A United States gallon of fresh water weighs $.33 lb. and contains 
231 cu. in. 

A cubic foot of water contains 1728 cu. in. or 7.48 U. S. gallons. 

A British Imperial gallon contains 277.20 cu. in., which is equivalent to 1.20 U. S. gallons, 
or 10 lb. in weight. 

Hie ncMinal pressure of the atmosphere is 14.7 lb. per sq. in.; it is equal to a column of 
water 34 ft. hi^ at ordinary temperatures. 

To find the capacity of a cylinder in gaUons, square the diameter in inches, multiply by the 
length in inches and divide by 294.1. 

To find the pressure in pounds per sq. in. of a column of water, at ordinary temperatures, 
multiply the height of the column in feet by 0.433. To find the head in feet, multiply the pres- 
sure in pounds by 2.31. 

*NoTB. U the netion and diaehu^gagM an not mt the Muselevtl, the total head mMt alto in^^ 

339 



340 



POWER PLANTS AND REFRIGERATION 



^ Total Head on Pump* Referring to Fig. 1, the velocity of flow from the rounded <»ifioe 
(A) at the base of the standpipe shown by dotted lines at the left <^ the figure, discharging uiuler 
a head H, will be the same according to Torricd li's Theorem* as that acquired by sbody faOiiig 

freely through the same height, or 9 « \2g H (velocity in ft. per sec.), H measured in feet. 

If the standpipe be attached to the suction pipe of the pump as shown, the pump phinger 
bdng held stationary, the velocity <^ discharge from the ddivery line at the reservoir eannoi 
be figured by the above formula, as a portion <^ the head H is balanced by the total meamired 




^ian^Pfpe 



TbM /fead • if0a(3Uf€i^Nea^ ^ Sum of lo3^ /fmag^ 

rfino/ ^eha/y/f&acf 



^Sfj/fip 



Fio. 1. 



head (At + h%) and another portion is required to overcome the frictional resistance in the punoip 
and line, the sum <^ the '' lost heads." 

The head lost between any two points is measured by the difference between the dynamic 
heads at the points considered. 

In pumping problems it is convenient to separate the suction head from the detivery head 
on account ^ the fact that the water is lifted to the pump by the atmospheric pressure acting 
on the surface <^ the water. The suction lift is consequently limi ted. 

Let H *^total head in It. required" fo" overcome "alT fnctional resistance, actually lift the 
water and create the velocity of discharge. 
h\ « head lost at entry to suction pipe. 
h\ « head lost in friction in suction pipe. 
ht « head lost in pump sucticm valves. 

Y • 
h% - velocity head in suction pipe - -r^, F, « veL in ft. per sec 

hj » measured suction head. 
Hg " total suction head. 

- Ai + A, + *, + ^ + ^ 
h4 « head lost at entrance to pump delivery and in deUvery valves. 
h% - head lost in friction in delivery pipe 



«1W a dlNUiiknof thii tlMoremieetliB Ctaapteron '* W»t«r, Steam and Air.' 



PUMPS 



341 




Swop 



FlO. 2. 



342 



POWER PLANTS AND REFRIGERATION 



At ~ final velocity head > 






V4 B veL in ft. per seo. 



h% « measured delivery head. 
Hd » total delivery head, ft. 

H » {Hf -^ h%) +^i>,^ssuming the miction and del ivery pi peg to be of t heaame 
= (*i + *f + ^ + ^ + W + (^7 + W +K 
» (lost heads) + (measured head) + (final velocity head). 
The total head to be overcome by the pump is therefore equal to the sum of the lost heads 
+ sum of measured heads + final velocity head. The final velocity head ht being small, is 
ordinarily neglected in calculations involving the flow of water. 

The head lost by friction in the pump suction valves, discharge valves and at entiy to the 
delivery pipe need not be considered in pumping problems as the sum of the heads lost throo^ 
the pump is taken care of and included in the efficiency factor of the pump. 
The total head for which a pump is selected is therefore: 

B » (^1 + As + Af) + (^7 + h%) feet or (sum of lost heads) + (sum of measured heads). 
Let hy s measured distance from center of pump (Fig. 2) to center of suction gage, feet. 
As » measured distance from center of pump to center of discharge gage, feet. 
Va ^ W¥B^ pressure lb. per sq. in. from miction gage reading (note that this will be be- 
low atmospheric pressure). 
Pd ^ S&ge pressure lb. per sq. in. from discharge gage reading. 
The total head against which the pump is operating is therefore: 

144p, ^'' . 144prf 4. iL . ^A 

where d is the density of the water under the given conditions. 

If the gage on the suction pipe is above the center line of the pump hy is minus, and if below, 
as shown in Fig. 2, ^ is plus. 

\/ Limit of Suction Lift. The atmospheric pressure (14.7 lb. per sq. in. absolute at sea level) 
will support a column of water (temperature 65°) h « 2.31 X 14.7 or 33.96 feet high. 

In <Mtler, however, to obtain the full effect of the atmospheric pressure it would be neces- 
sary to create a perfect vacuum on top of the water column. This is an impossibility owing 
to imperfections in the pump and from the fact that a vapor tensicHi exists over the surface of 
the water corresponding to its temperature. The vapor tension in lb. per sq. in. absolute pr«B- 
sure corresponding to various temperatures is found in the steam tables from which the correspond- 
ing heights of water equivalent to the pressures are readily calculated. Table 1 was calculated 
in this manner. 

^ TABLE 1 

MAXIBHTM THEORETICAL HEIGHT TO WHICH A PUMP CAN LIFT WATER BY SUCTION AT 

DIFFERENT TEMPERATURES 

(Barometer 29.92) 





Temperature of 


Maximum Tbeoretieal 


Temperature of 






Lift. Feet 


Water 'F. 


Uft,F^et 




40 


88.6 


180 


29.2 




60* 


88.5 


140 


27.8 




60 


88.4 


150 


25.4 




70 


88.1 


160 


28.5 




SO 


82.8 


170 


20.8 




90 


82.4 


180 


16.7 




too 


81.9 


190 


12.8 




110 


81.8 


200 


7.6 




120 


80.8 


210 


1.8 



PUMPS 



343 



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POWER PLANTS AND JUSFRIGERATION 



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PUMPS 



345 



In i^aciice these lifts cannot be obtained; 20 to 25 ft. measured suction head is considered 
a practical limit in pumping water at ordinary temperatures. (60 to 70 degs. F.) 

If air be admitted to the suction line near the surface of the water, the column becomes a 
mixture of water and air, and owing to the decreased density of the mixture the limit of lift is 
greatly increased. This scheme is, however, not considered very practical and is not resorted to 
except in cases of emergency. In pmnping hot water it is always advisable to have the water flow 
inio the sueiion chamber by gravity. This statement applies particularly to boiler-feed pumps 
drawing their supply from feed-water heaters. 

Head Lost by Entry to a Pipe. The head lost at entry to a straight pipe is usually stated 

as ^ » 0.6 TT » ^ ■ velocity in ft. per sec. 

i/Head Lost tfaxoagh Pomp. The head lost through the pump valves and passages is difficult 
to estimate and naturally varies with the construction. This loss of head, however, does not 
eater into the calculations, as it is included in the efficiency factor (e) of the pump and need not, 
therefore, be considered. 

It is assumed that the efficiency referred to for a reciprocating pump has been calculated 
by using the delivered water horsepower, as calculated from the gage readings, and not the 
indicated horsepower of the water end as determined by means of an indicator. The latter if 
used in calculating the efficiency gives the mechanical efficiency of the pmnp and includes the 
loss occasioned by the friction cA water through the pump 
>/Head Lost by Pipe Friction* 

h » head lost in ft. 

L « length <^ pipe in ft. 

D -> internal diam. of pipe in ft. 

V « velocity water. 

/ « coef . friction. 

L y 

Fanning gives for clean pipe the following average values (Table 4) of / for velocities of 1.7 
to 7 ft. per second. These coefficients vary considerably with difiPerent authorities. 

Fen* additional data on the flow of water and friction pressure loss chart, see Chapter II, 
Table 3, and Fig. 35 at the end of this chapter. 

TABLE 4 





Diametar Pipe, Inches 


Valub of / 




100 Feet 
per Minute 


400 Feet 
per Minute 


1 


0.034 
.080 
.027 
.025 
.024 


0.026 


s 


.024 


6 


.022 


10 


.021 


M 


.020 







^ead Lost by Friction in Elbows. The following formula by Weisbach is commonly used 
to approximate the head lost through ells: 



k, . [o.l31 + 1.847 (-^)"] X ^ X 



a 
180 



in which r » internal radius of pipe in feet, R « radius of curvature of axis <^ pipe, F** velocity 
in feet per sec., and a « the central angle or angle subtended by the bend. 



346 



POWER PLANTS AND REFRIGERATION 



TABLE 5 

^ LOSS OP HEAD IN 90* BENDS AND ENTRANCE HEAD IN FEET FOR VELOCITIES OP 1 TO 15 

PER SECOND 



Vdodty, Pert 
per Second 



of 
Head 

in 
Peet 



— - 1 



f 
~R 

r 
R 



0.0160.061 



EntrmnoeHead 
in Peet 



0.01 



2 



0.1880.2470.8840.6660.7680.986 



0.08 



8 



— - VJO.0080. 0090.0210.0860.067*0.0820. 11 



0.07 



0.18 



0.196k). 28 



6 



0.88 



8 



0.16 



0.60 



8.6 



1.111.261.861.68 



9 



9.6 



10 



10.6 



0.670.680.700.780.860.981.02 



1.69^1.802.02^. 



11 



11.6 



. 17|0 . 19|0 .21|0 .28)0 .26|0 .27)0 .80j0 .88)0 .8810 .46)0 .61 

1.75 



12 



222.608.028.47 



18 



14 



1.121.81 



1.51 



16 



The accuracy of this formula when applied to standard fittinjgs is questionable; the usual 
method employed to allow for the friction of ells and valves is to add to the measured length 
of the line various amounts as indicated in Table 6. 

One prominent maker of hydraulic machinery makes use of the following data in estimating 
the friction head of fittings. The following equivalent length in feet <^ straight pipe should be 
added for each fitting in figuring friction. 

TABLE 6 

^ PRICTION OP STANDARD PIPE PITTINGS 

Equivalent length of straight pipe to be added to meaaiared length 



Siae of Pitting 


K" 


1" 


lyr 


IJi" 


2" 


2>i" 


8" 


4" 


6" 


€r 






Elbowa 


6 

10 

6 


6 

10 

6 


6 
12 

7 


7 

14 

8 


7 

16 

8 


10 
20 
12 


12 
24 
24 


18 
86 
80 


26 
60 
40 


SO 


Return benda. 


60 


Globe valvea 


50 







Owing to the burr (caused by cutting the pipe with a ^dieel cutter) obstructing the flow in 
the smaller pipes, it is advisable, unless the burrs are reamed out, to multiply the above figures 
by 3 for ^-inch and I-inch fittings and by 2 for l}4t 13^ &nd 2-inch fittings. 

TABLE 7 

COMPARATIVE LOSS OP HEAD IN PITTINGS AND VALVES 
^ (Experimenta of John R. Freeman) 



Name of Pitting 


Number of Peet 

of Clean. 

Straight Pipe 

of SameSiae 

which would 

Cauae the Same 

Loaa aa Pitting 


Name of Pitting 


Number of Feet 

of dean* 

Straight Pipe 

of SameSiae 

which woidd 

Cause the Same 

Loaa as PlttlBg 


6-in. Swing diedc yalve 


60 
200 

26 
180 

4 
9 
9 


8-in. to 8-in. ahort-turn teea 


17 


6-in. Lift dieek valve 


Hbend 


6 


4-in. Swing check valve 


6-in. Grinnd dry i^pe valve 


80 


4-in. Lift ebeck valve 


4-in. Grinnd dry oipe valve 


47 


2 V4-in. to 8-in. long-turn dla 


64n. Grinnd alarm cheek valve 

44n. Grinnd alarm check valve 


100 


2 Vi-in. to 8-in. ahort-turn ella 

8-in. to 8-in. long-turn teea 


47 







^ Allowable Velocity of Water tiirough Pipes. From the diagram <^ friction head of pipee, 
Fig. 6, Chapter II, and Table 3, it ia seen that the head lost increases very rapidly as the 



PUMPS 



347 



velocity and qoantity discharged increases. In order to prevent excessive loss, practice has shown 
that the velocity of water in the discharge line should ordinarily not exceed 360 to 4S0 ft. per min. 
and approximately 200 ft. in the suction line for reciprocating pumps. Suction and discharge lines 
for ordinary length of runs are usually made the same size as called for by the flanges on the 
pomp, never smaller, and preferably larger when the runs are long, particularly the suction line. 

The practice of centrifugal pump manufacturers is to allow a discharge velocity of 10 to 
12 ft. per sec. or 600 to 720 ft. per min. at rated capacity of pump. 

The e&ae o£ outlet of a centrifugal pump, however, is no gage whatever of the proper size of 
piping to attach to it. 

v'DIRECT-ACTING STEAM PUMPS 

A steam actuated pimip without a flywheel is known as a direct-acting steam pump. This 
type of pump having comperatively few working parts requires little attention and is in gen- 
eral very reliable. For general service about a power plant — boiler feeding, vacuum pumps, 
etc. — ^it is the most popular of all types. It is built either simplex or duplex. 

i/^mplez Pump. This type is built with one steam cylinder and one water cylinder. It 



I' 







Steam £nd 

Fio. 3. Simplex Pxtmp. 



employs a steam-'thrown main valve in order to obtain a reversal of stroke, the action of which 
will be understood by reference to Fig. 3. The sketch refers to no particular make of pump 
and is intended to illustrate only the principle involved. 

Just before the piston reaches the end of its travel the auxiliary steam valve, moved in the 
opposite direction by the link motion connected to the piston rod, uncovers the small steam port 
whidi admits steam back of the piston a, which in turn operates the main valve h. The 
movement of the main valve uncovers the steam port connected with the end of the cylinder 
toward which the piston is moving and at the same time uncovers the exhaust port at the other 
end* causing a revosal of the moticMi of the piston and plunger. 



DUFI.EX Steam Puup. 



FlQ. 6. Ddplkx Pdmp Valvb Gbak. 



PUMPS S49 

Seconunended capadtiee for Torious sJiea of aimplei putnpd are giveil by TUiIe S. 

Ihiplex StMm Pump (F^. 4 and 5). This type of pump ia built with two 8(«ain cylinden 
mod two water cylinden; that ia, two simplex steam pumps placed aide by aids with a valve 
motion so designed that the movement of the slide valve is controlled and operated by the opposite 
pump. As one piston moves to the end of its stroke and is gradually brought to rest, it moves the 
slide valve of the opposite steam cylinder admitting steam back of the piston which is at reat, 
causing it to similarly move forward to the opposite end of its stroke. 

The valves have neither " lap " nor " lead," and there is anme lost moticoi allowed between 
the valve and the ofterating mechanism, which causes the piston to pause at the completion cd 



Ci;sHioH Port. 

the stroke. The pistons are in action or moving approximately only J^ of the time. The cyl- 
inders have five ports. The two end ones are for the adiiuaaion of steam and the three inner 
ones for the exhaust. The piston is cushioned at the end of its stroke and prevented from strik' 
ing the cylinder head by the piston passing over the exhaust port in the steam cylinder and trap- 
ping the steam in the cleanwoe space between the piston and the cylinder head. 

Fm steam cylinders 14" diam. and above on auxiliary port is ordinarily provided uid ia 
placed beyond the regular exhaust port and communicating with it through a valve (Fig. 6). 
TiuB arrangement allows the cushicxiing action and the length of stroke, to some extent, to be 
ngulat«d. A 12" stroke pump has a piston clearance of approximately )^" at each end of the 
oylindo'. 

Recommended capacities for various sizes of duplex pumps are given by Table 8, and di- 
mensions by Tables 10 and 11. 

VanouB typCB of water ends used for direct-acting steam pumps are shown by Figs. 7, 8, 
and 9. 

One of the 'advantages of the qenter and outside packed plunger is due to the fact that any 
leakage past the plunger may be immediately detected and readily stopped by an adjustment 
of the packing from the outside, whereas with the piston type it is necessary to remove the pistcm 
from the rod when repacking is necessary. In order to even up the velocity of discharge an air 
chamber should always be provided on the discharge side in case the discharge pipe is bng. 

^Power Pumps. In this class of pnmp (Pig. 31) the water plungeis are actuated by means 
ol a crank and connecting rod. A gear wheel mounted on the crankshaft moshM with a pinion 
on the receiving pulley shaft. 

The most popular type, for general service, is constructed with three vertical single-acting 
cylinders and termed a triplex power pump. The cranks are placed 120° apart, which arrange- 
ment gives a fairly steady discharge. 

The pump is genially driven by means of an electric or an internal combustion motor 
or iteam en^ne through a belt or ulent chain, although sometimes geared direct to the 
armature shaft with a rawhide pinion. 

llus type of pump is largely used when electric power is available for all classes of service 
>* boiler feeding, house service, elevator service, hydraulic press-work, etc. See Table 20 lor 
qieedi and capacities. 



350 POWER PLANTS AND REFRIGERATION 

For intermittent house-tank service the motor is controlled by means of a float in the 
tank which operates an automatic switch, o(mtrolling the motor, as the water levd in the tank 
rises or falls. 

The efficiency (water hp./brake hp.) of this class of geared pump is approximatdy 75% 
(Fig. 10). When a power pump is used f<n' boiler feeding and driven by a motor the rate of 
pumping may be varied by controlling a by-pass valve located between the suction and discharge 
side of the punq). 

Reciprocating Pump Capacity. 

Let Q >■ cu. ft. per min. actually required to be pumped. " 
D >■ cu. ft. per min. plunger displacement. 
E *■ volumetric efficiency of water end. 

— -rr and D — -=: 
D E 

d *■ density of water corresponding to its temperature (62.3 lb. per cu. ft. at 60*). 
h >■ diameter vrater cylinder, inches. 
8 *■ stroke, inches. 

n « number of working strokes per min., each vrater cylinder. 
Plunger displacement per stroke cu. in. » 0.7854M^. 



Plunger displacement cu. ft. per min. 



0,7S5WSn 
1728 



» 0.000454&>5n. 
Plunger displacement in lb. per min. » D X <i. 
Plunger displacement in gal. per min. b 7.48D. 
Diameter water end in inches, 



A0.7854Sn A 0.7854n^5 * \ nE. 



nES 

*^ Volumetric Efficiency of Reciprocating Pumps (E). The voliunetric efficiency of a recipro- 
cating pump is defined as the ratio of the volume actually discharged to the plunger displacement 
in a unit of time. 

Owing to the fact that it is impossible to absolutely prevent leakage by the valves and piston 
or plunger of a reciprocating pump the actual displacement of the plunger must be greater than 
tike quantity of water to be handled. Ebqperiments conducted by the Inspection' Department of 
the Aaaociaied Factory MtUtuU Fire Insurance Cos. on a number of duplex steam pumps show that 
in a new pump with clean valves and air-tight suction pipe and less than 15 ft. suction lift the 
actual delivery is only 1)^ to 5% less than the plunger displacement. As the slip increases with 
wear 10% may be considered a fair allowance to cover slip, valve leakage, etc. Tlie value of E 
may cnxiinarily be assumed in calculations as 85 to 00% with pump in fair working condition. 
The rated capacity of pumps as given by manufacturers' catalogs refer to plunger diq)laoementy 
consequently a deduction of approximately 10 to 15% from the capacity stated should be made 
to cover shp and leakage. 

' Number of Working Strokes per Minute (n). In order to reduce to a safe margin the strains 
and consequent wear on the working parts of the water end, which are produced mainly by 
the impact of the piston or plunger on the water, the number of reversals or strokes per min. 
must necessarily be limited. The custom of rating pumps at a piston speed of 100 ft. per min. 
is becoming obsolete. Experience has shown that for long life and good soince pumps which 
are to operate continuously as for boiler feeding, water works, etc., the number of working strokes 
per min. should not (Htiinarily exceed the values given in Table 8. If the pump is to be only 
occasionally operated as in the case of fire pumps the speed may be practically double the tabular 



: n 





Fzo. 7. Packxd Pibion. 



FlO. 8. OBMTBB OUTBIDB PAGKZNa. 




Pio. 9. Oxttbids-Packsd Plunobrs. Ttpbs of Watbb Bndb fob Stbam Pumps. 




60 76 100 125 

ToM H994j Lb, Per Sq, Ui, Giuge 
8pe^d GojiitsiiL Htsd YtrJMj^tt 

Fsa. 10. 



352 



POWER PLANTS AND REFRIGERATION 



values as the consideration of wear for the brief periods during which the pun^ is operated 
not an important item. 

.^TABLE 8 

CAPACITIES OF BOILER FEED PUMPS 

Sini^: Steam Actumted 
Duplez: Steam Actumted 
Triplez: Power 





Pounds 


Gallons 


Single Pump 

1 


DupuBX Pump 


Tmpuek Pump 


Hp.of 
BoUer 


















Hour 


Minute 


Sbe 


No. 
Stka. 




Siae 


- 


No. 
Stka. 


Siae 


R.P. 
M. 










Min. 






. 


Min. 




60 


1880 
2742 
8665 


8.8 
6.7 
7.6 




* • 

• • 

27 


8 


z 2 
ix 8 
Ix 8 


z 4 
z 4 
z 4 


40 
27 
86 


2x8 

2Hx 4 
2Hz 4 


82 


76 




24 


100 


6Hx 8zl0 


81 


160 


6489 


11.4 


6Hz 8zl0 


41 


Bh 


Ix 8^ 


\x 6 


81 


8 z 4 


82 


200......... 


7820 


16.26 


6Hz 4zl0 


80 


6^ 


ix Zh 


z 6 


41 


4 z 4 


24 


260 


9162 


19. 


6Hz 4zl0 


88 


6^z Bh 


\x 5 


61 


4 z 4 


29 


800 


10990 


22.9 


6V|z 4zl0 


46 


6 


X 4 


z 6 


89 


4 z 4 


86 


860 


12820 
14662 


27. 


IHx 6zl2 


29 


6 


z 4 


z 6 


42 


4 z 6 


28 


400 


80. 


7Hz 6zl2 


88 


7 


z 4 


zlO 


81 


4 z 6 


81 


600 


18810 


88. 


7Hz 6zl2 


41 


7 


z 4 


zlO 


41 


4 z 6 


88 


600 


21976 


46. 


9Hx 6zl2 
9Hz 6zl2 


84 


7 


X 4K 


izlO 


88 


6Mz 6 


2S 


700 


26642 


58. 


40 


7 


z 4^zl0 


44 


6Hz 6 ' 


80 


800 


29300 


61. 


11 z 7zl4 


28 


8 


z 6 


ZlO 


40 


6Uz 8 


25 


1000 


86625 


76.2 


11 z 7zl4 


85 


10 


z 6 


zl2 


28 


6Hx 8 


82 


1200 


43956 


92. 


11 z 7zl4 


45 


10 


z 6 


zl2 


86 


7 z 8 


24 


1600 


64940 


114. 


IBHx 8zl4 


41 


10 


z 6 


zl2 


44 


7 z 8 


29 


1800 


66980 


137.4 


18Hz 8zl4 


47 


12 


z 7 


zl2 


88 


7 z 8 


86 


2000 


73269 


162.6 


ISUz 9zl8 
18Hz 9zl8 


88 


12 


z 8 


zl2 


83 


8 z 8 


80 


2600 


91602 


191. 


41 


12 


z 8 


zl2 


40 


8 ZlO 


80 


8000 


109890 


229. 


16 z 10 z 18 


41 


12 


z 9 


zl2 


88 


8 ZlO 


86 


8600 


128200 


267. 


18 z 12 z 24 


25 


14 


zlO 


zl2 


41 


9 ZlO 


84 


4000 


146626 


806. 


18 z 12 z 24 


28 


16 


zll 


zl2 


88 


9 zl2 


81 


4600 


164880 


843. 


20 z 14 z 24 


24 


16 


zll 


zl2 


88 


9 zl2 


86 


6000 


183166 


382. 


20 z 14 z 24 


27 


16 


zl2 


zl2 


86 


10 zl2 


82 


6000 


219778 


468. 


20 z 16 z 24 


26 


16 


zl2 


zl2 


48 


11 zl2 


81 


7000 


266410 
298044 
866800 


634. 
610. 
764. 




• ■ 

• • 

• • 


20 
20 
20 


zl4 

zl4 
zl6 


zl6 
zl6 
zl5 


29 
88 
86 


12 zl2 
18 zl2 
14 zl5 


81 


8000 




80 


10000 




26 









"^ Direct-Acting Simplex or Duplex Boiler Feed Pomps. The boiler feed pump should be 
large enough to take care of the maximum overload of the boilers without having to operate 
at an excessive speed. The speeds specified in Tables 8 and 12 for various lengths of sttoke 
will give satisfactory service. Boiler feed pumps should be installed in duplicate or at least in 
conjunction with an injector to prevent a shutdown of the plant in case this important auxiliary- 
is put out of commission by some cause. 

A boiler hp. being the evaporation of 34.5 lb. water from and at 212** per hr., the actual 
pounds of water required per hr. for various combinations of feed water temperature and boiler 
pressure is found by dividing 34.5 by the factor of evaporation corresponding to the actual con- 
ditions of operation. In approximate calculations 30 lb. per actual boiler horsepower may be 
assumed. 

The steam consumption of simple direct-acting steam pumps, taking steam full stroke with- 
out expansion, ranges from 100 to 175 lb. per indicated horsepower-hour of steam end, and f<Nr 
the compound duplex approximately 50 to 100 lb. These figures are great^ exceeded by pumps 
in poor condition. 

As the exhaust steam is almost invariably utilized in a feed water heater there is frequently 
no necessity for using an economical feed pump, except in cases whare there is an exoeas of ex- 
haust steam from the main units available for this purpose. 



PUMPS 



353 



TABLE 9 

STEAM PUMP TESTS 
(Robert Huni A Comptm/f) 





Siie,IiidiM 


StrokM 
Mkute 


Gmilon 
Actual 


Total 
Head 


% 
SUp^ 


Effi- 
dencyt 


SiBAM Consumption 
PBR Hour 


S^yle of Pump 


I.Hp. 

Steam 

CyUnder 


Ddiverad 
W.Hp.t 


DoploK 


6z4z 6 
6z4z 8 
8 z 6 z 12 
8 z 6 z 12 


79. 

64.7 
164.5 
127.7 


20. 

24. 
189.8 
185.6 


191. 
199. 
288.6 
241.2 


16.16 
7.12 
1.28 
8.84 


0.78 
0.90 

• • • • 
■ • • • 


171.2 
182.4 


196.8 


Smplez 


146.8 




82.67 


Sbnpimx, ..'.'.'.'.'.'.'.'.'.'.'.'. . 


111.74 







* Par eent loaa due to ■Up and diort itrokaa. 

t Tho eAciency here referred to ie the ratio deUvered honepower/indicated boraepower of steam cylinder. 
It inctndea all loaaee througfa the pomp. 

t Thia waa calculated from the total steam consumption per hour and the hp. determined from the total head 
as read bar the gagea and the actual weight of water pumped. 



The steam consumption of direct-acting pumps decreases as the speed is increased. A test 
on a 16" X 12'^ duplex steam pump at the Mass. Inst, of Tech. gave approximately the following 
results: 

20 strokes per min., 200 lb. per 1 hp.-hour of steam end. 

35 strokes per min., 150 lb. per 1 hp.-hour of steam end. 

75 strokes per min., 100 lb. per 1 hp.-hour of steam end. 

Allow a drop in steam pressure between the boiler and pump of at least 5 lb. per sq. in. in 
checking for size of steam cylinder of pump. The ratio of the area of steam cylinder to the area 
of the plunger varies from 2 to 3, so that cnxiinarily no calculation is necessary for the steam 
end for boiler feed pumps. 

''lH>wer Required to Raise Water. 
Let Q » cu. ft. per min. 
H » total head in ft. 
d ■■ density of water. 
Then ddivered water hcnvepower is 

QHd 



w.hp. 



33000 



The power required to be applied at the pulley brake hp. of a power pump or a centrifugal 
pump is found by dividing the w.hp. by the mechanical efficiency of the pump, approximately 
75% for power pumps and 65% for ordinary centrifugal pumps (Fig. 11 and Table 13). 

The i.hp. of the steam end for the flywheel type of pump is approximately equal to the 
i.hp. water end divided by 0.80. 

• 

Percentage of Steam Generated by Boilers Used by Direct-Acting Feed Pumps. Let 
I.hp. B Ind. hp. of main engine. 
i.hp. a Ind. hp. of pump steam end. 

e « efficiency «■ delivered water hp./i.hp. of steam end. 

Delivered water horsepower 



iJip. 

W 

w 
d 



eff. of pump 

steam consumption main engine per I.hp.-hr. 
steam consumption pump per i.hp.-hr. 
density of feed water. 



364 POWER PLANTS AND REFRIGERATION 

Q >■ cu. ft. per min. pump must Biq)ply boiler 
I.hp. X TT + i.hp. X w 

eoxd 

D >■ pump displacement cu. ft. required per min. 
Q _ I.hp. X W + Lhp. X w 

" e" eoxd x^ 

Pi >■ boiler pressure in lb. per sq. in. gage. 
The friction pressure loss and measured head, in lb. per sq. in., of feed Une may be assumed 
as 0.25 Ply when the layout of the lines is not g;iveD. Based on this assumption, 144(1.25 pi) » 
lb. per sq. ft. pump must operate against. 

_ 144(1.25 p,) XD 

*• ^' " 33000 X e 

Substitute value of D and solve for the indicated hors^iower of steam end. 

^ Pi X lhp. X W 

** ^' " 11000 XeXdX^-piW 

100 i.hp. X w 
Per cent total steam generated used by feed pump = 77 =— ; — — . 

^ Rzample. Non-condensing plant with boiler feed pump as the only auxiliaiy. 
Boiler pressure pi -* 100 lb. gage. 
I.hp. of non-oondensing engine <■ 100. 
Steam consumption of engine <■ 32 lb. i>er I.hp.-hr. 
Steam consumption of pump <■ 125 lb. per Lhp.-hr. 

Required the indicated horsepower of steam end of pump and the peroentage of the total steam 
generated that will be used by the pump, e » 70%; E -* 85%; d <- 62.4, correqMnding to a feed water 
temperature of 54** F. 

^ 100 X 100 X 32 

« 0.81 
Peroentage of total steam generated by boilers used by the feed pump 

0.81X125 ^^^^ ^^^ 

•- 0.031 or 3.1%. 

100X32+0.81X125 ^^ 

^ Size of Steam Cylinders; Direct-Acting Steam Pumps. 
Let i7 « total head in ft. 
d » density of ¥rater. 
Pw >■ theoretical effective pressure, lb. per sq. in. on water plunger or piston. 

Ed 

" 144 

A " area plunger sq. in. 

a « area steam piston sq. in. 

p « initial steam pressure at pump, lb. per sq. in. absolute, 
pi " back pressure absolute, (nrdinarily assumed 15.7 to 16.7 lb., atmospheric ezhaost. 

delivered water hp. _ . 

e - TT TZ zr-, = pump efficiency. 

i.hp. of steam cylinders 

The delivered hcn'sepower is equal to the totoZ htad pumped against X weigjht of water pun4>ed 
per min./ 33,000. 

PvA " a(p — pt}e (Plunger thrust » Steam piston thrust X ^)- 

(p - pi).e 



It IB advisable to allow & drop in the steam pressure between boilo' and steam cylinder erf 
apprcnc. 5 to 10 lb. per aq. in. 

ConqMuad Direct-Actliig Steam Pumps. In this type of diiect-acting pump st«am is tA- 



Section of Compound Duplex. Pump tw»eA»(j» 







i 



=^ 



n Pim. Uot/- 

Staam End \ 



r^ 



Canter Qutsidt Packed 
^ater £nd ^ 



nutted to the high-pre»ure cylinder lot the full stroke and exhausted tar the full strcJce directly 
into the low-pressure cylinder. 

The steam is worked expanmvely, due to the fact that the volume of the low-preaaure cylinder 
is mevtseal times that of the high-pressure cylinder, the total number of expanaionB being equal 
to the ratio, vol lov-preasure oylinder/voL of bigb-iweeaure cylinder. 



356 POWER PLANTS AND REFRIGERATION 

Due to the fact that the steam is worked expansively the steam consumption of the oompouod 
type is only about one half that of the simple type and when an economical direct-acting steam 
pump is desirable this t3rpe of pump is frequently used. A test on a 10" X 16" X W X 16" 
compound direct-acting pump, reported by Robl, HitrU and Company , gave the following results: 
Steam consumption (djy) 45.56 lb. per delivered water hor8eix>wer-hour. Effidoicy 92%. 
SUp2.53%. 

The steam thrust of a compound pump is variable, due to the fact that the back p r cmu re 
on the high^ressure piston and the forward pressure on the low-pressure piston is varying. As 
the pump is not equipped with a fl3rwheel, the maximum thrust on the plunger must always be 
less than the minimum combined thrust of the high- and low-pressure pistons. 

In order to determine the size of steam cylinders required for a given head and initial steam 
pressure the following formula may be used: 
Referring to Pig. 12. 

Let p « absolute initial pressure in HP. cylinder. 
Pi = absolute initial pressure in LP. cyUnder. 
pt = absolute terminal pressure in LP. cylinder. 
Pi s absolute back pressure in LP. cylinder. . 
Vhp, = volmne of high-press, cyl. in cu. ft. 
Vlp, = volume of low-press, cyl. in cu. ft. 

Vr = volume of receiver. 
Cjfp. = clearance volume HP. cyl. in. cu. ft. 
Clp, = clearance volume LP. cyl. in cu. ft. 

Vhp 

• = i2 = cylinder ratio. 



Vlp. 

The following assumptions may be made in approximate direct-acting compound steam 
pump calculations: 

Cbp. = 0,20Vhp.; Clp, = 0.20 Vu>.; Vji^ Vhp. 

To simplify the computations the following substitutions may be made: 
Let Vhp, = 1| then 

Vr = 1, Chp, - 0.20, Vhp.^ RVlp.. Clp, = 0.20^ 

In a problem of this kind p and pi are always known or may be assumed with reason- 
able acciu^cy. At the beginning of the stroke (Fig. 12) the pressure acting on the front face of 
the HP. piston is p, that acting on the back face p\. The pressures acting on the low-pressure 
piston are pi and p%. At the end of the stroke these pressures are p and pt and pt and pi. Of 
these we know p and p%. It will be necessary to solve for pi and pt. 

Consider that the pistons have completed a stroke to the right. The high-pressure cylinder 
to the left of the piston as well as the clearance on that side are filled with steam at pressure p. 
The valve has been shifted and escape of the steam to the receiver is cut off. Up to Uie instant 
the valve was shifted the receiver was in communication with the volume of the low-pressure 
cylinder to the left of the piston. The pressure in the receiver is ps, that being the final preasure 
after expansion in the low-pressure cylinder. The clearance space of the low-pressure piston 
to the right of the piston is filled with steam from the exhaust stroke just completed at a 
pressure pt. 

In the design of steam engines it is generally assumed that the relative changes of pressure 
and volume are related according to the equation po = const. It will also be assumed that 
when a number of chambers of different volumes and different pressures are put in communi- 
cation the resulting pressure will agree with the statement, 

pV + p"»" + P"V" = pr{v' -h »" + »"0 



PUMPS 357 

AflBume Uiat the valve Is shifted, placiiig the high pressure cylinder, the receiver and the 
dearaiioe q>aoe oC the low-pressure cylinder in communication. The following equation will hold : 

P(V'HP + CHp)+PfFii + P«CLF-pi(FHP + CHP+ Vr-^-Clp) (1) 

The valve remaining in the same position, the pistons now begin to move toward the left. 
The steam is being crowded out of the high^ressure cylinder into the receiver and the low-^res- 
Bure cylinder. Because of the sise of the low^^ressure cylinder there is an increase in volume 
and a decrease in pressure. The new equation is: 






(2) 



Substitutiiig the value of ps in equation (1) pi is obtained. Substituting pi in equation 

(2) p9 is obtained. 
The thecxetical thrust at the beginning and at the end of the stroke may now be obtained. 
Let Gap B aiea of high-pressure piston in sq. in. 

a^p >" area of low^ressure piston in sq. in. 
Then the steam thrust at the beginning of the stroke is 

anpiP" Pi) + ^LpiPi - Pi) (3) 

and the steam thrust at the end of the stroke is 

Ohp(P - Pt) + aij,(pt - Pi) (4) 

If a 15% loss is assumed then 85% of (4) will give the maximum permissible thrust on 
the water piston or plunger. 

(/Siample. Let it be required to determine the total head H against which a IC (HP dia.) X 16'' 
(LP dia.) X 10" (dia. plunger) X 16'' (stroke) compound direct-acting pump will operate for the fol- 

Vlp 16« 



lowing conditions: p - 116, pi - 14.7, Cap- 0.20 Fjfp. Clp -0.20 Vw* Vr =• Vhp, 



Vhp 10» 



2J56. Substituting the above values in equation (l) in terms of the high-pressure cylinder volume 

Vhp ^ 1, Vlp - 2.66, Clp - 0.512, Vk - 1 

116 (1. + 0.20) + 0.636 pi X 1 + U.7 X 0.6l2 - pi (1. + 0.20 + 1. + 0.612) 

Pi - 70.6' 



. A.+0.20-H.+0.812 \ _ 

^ V0.2O+L +2.56 +0.612/ 



Theoretical steam thrust at beginning of stroke is: 

78.64 (116 - 70.6) + 201.06 (70.6 - 14.7) - 14,806 lb. 

Theoreticai steam thrust at end of stroke is: 

78.64 (116 - 44.8) + 201.06 (44.8 - 14.7) - 11.644 lb. 

The expected or actual effective minimum thrust at water end will be: 

0.85 X 11,644 « 9897 lb. 

9897 

H - 291 ft. 

78.64 X0.433 

Duty. The term duty is often used as an efficiency standard in connection with steam-driven 
pumping machinery. It refers to the number of ft.-lb. of delivered work done by the 
pump for a certain quantity of heat energy supplied. Duty may be defined either as the number 
of ft.-]b. of delivereid work done in lifting the liquid per 1,000,000 B.t.u. used in driving the 
pomp; or it may be defined as the number of ft.-lb. of work per 1,0(X) lb. of dry steam used. 



358 POWER PLANTS AND REFRIGERATION 

The firat form is the better. Duty is then equal to— 

v>(xiri + gi - St) 

in whidi W = \b.ot water pumped per min. 

H = actual total head on the pump in ft. 
to — lb. of Htcam ueed per hr. 
xin +gi — heat content above 32° F. per lb. at steam supplied. 

9i — the heat <rf the liquid per lb. of steam at the praesure in the exhauat. 



TABLE 10 

DIUENBIONS or DEAN BROS. DUPLEX FUHP8 
(OuUld* Padnd Plimt«n) 



Hi 10 



lis 



iK 



TABLE II 

DIMENSIONS OP DEAN fiROS. DUPLEX FUUFS 



iilHxB BMiSKiG Si4ii 



7 



i CENTRIFUGAL PUMPS 

Aa tbe name implies, the preasure generated by the pump is due, largely, to the action of 
centrifuga] force imparted to the water by means of a bladed impeller rotated in an enclosed 
caaing. The cenbifugal pump aa now coDstmcted in its several forms is adapted to practically 
&11 claaaes d pumping service for which reciprocating pumps are used with the poesible exception 
ri hi^-presBure hydraulic press and similar service. 

It is particularly well adapted for direct oonnection to engines, turbines and motors. It 
'lias only one moving part, the impeller, no valves to get out <A order, and 'u therefcse subject 



360 



POWER PLANTS AND REFRIGERATION 



TABLE 12 

DUPLEX STEAM DIRECT-ACTING BOILER FEED PUBiPS 

Cftpaeities and 







Actual 










Recom- 


DeUvoies 


Capacity 






Siae 


mended 


Baaed on 


Boilers 




Net 


Diameter Steam Cylinders 


Strokes 


80 Per Cent 


Pump 
Wilf 


Weight. 


Prkse 


Diameter Water Cylinder z 


Mkute 


Volumetric 


Pounds 


at 


Stroke, Inches 


Efficiency, 


Serve, 




Factory 






Pounds 


Horsepower 










per Hour 









Piston Pattern 



4Hx 
6Hx 



8 x2 X 4 
8x4 
8Hx 6 

6 x4 X 6 

7 x4Hx 7 
7x6x7 
7 x4HzlO 

7 X 6 X 10 

8 X 6 X 12 
10 X 6 X 12 



60 
60 
46 
46 
40 
40 
40 
40 
40 
40 



2,174 
4,890 
7,487 
11.787 
16,406 
19,018 
22,006 
27,168 
82,602 
46,946 



60 
100 
160 
260 
826 
400 
460 
676 
676 
1,000 



210 

860 

490 

640 

990 

1.026 

1.876 

1.876 

1,610 

2,476 



$81 

61 

69 

80 

112 

112 

168 

161 

173 

260 



Plunger Pattern 
Outside Cent«r Packed 



4Hz8 X 4 

: 6Hz8Hx 6 

6 x4 X 6 

7Hx6 X 6 

7Mz4^xlO 

10 X 6 X 10 

12 X 7 X 10 



60 
46 
46 
45 
40 
40 
40 



4,890 
7.487 
11.787 
18.888 
22.006 
89.122 
68,260 



100 
160 
260 
400 
460 
800 
1,100 



880 
616 
726 
1,800 
2,400 
8,200 
4,000 



$69 
88 

108 
187 
256 
844 
440 



Note. — ^If pump is to be brass fitted add 16 per cent to above prices. 



to little depreciation. The following; facts, however, should be borne in mind, when selecting 
a centrifugal pump. The efficiency of a centrifugal pump is quite a variable quantity, depending 
primarily upon the head against which it operates. Centrifugal pumps are rated by the manu- 
facturer at or near their maximum efficiency at a definite speed at which there is but one 
head under which the pump will operate at maximum efficiency. Therefore if the pump is 
expected to deliver the tabulated quantity of water at the speed and the power consumption 
stated, the conditions must be reproduced in practice under which the pump was originally 
rated. 

The above mentioned points will be apparent from an inspection of the characteristic curves 
of a centrifugal pump. Fig. 15. This diagram shows the relation between the capacity, head 
and horsepower for a fixed speed. The rated speed having been chosen after a series of tests 
run at various speeds to determine the speed which gives the highest efficiency. 

It is essential, to the proper selection of a centrifugal pump and its drive, that the method 
of rating this t3q)e of pump be clearly understood. 

Attention is directed to the comparatively low efficiency of the smaller sizes of centrifugal 
pumps, and such sizes are to be avoided if a reciprocating pump may be used for the purpose at 
hand. 

Ni Efficiency of Centrifugal Pumps. The efficiency (water horsepower output/brake horse- 
power input) naturally varies somewhat with the dififerent designs of various makers. The fol- 
lowing table of efficiencies compiled by Af . W, Ehrlich and published in the " Practical Engineer," 
September 1, 1915, is the result of an extended study of the subject and is based on the averages 
of a large number of tests of various makes, including both the volute and turbine types, when 
operated at their rated capacity. 



PUMPS 



361 



TABLE 13 

PUMP SIZES, CAPACITIES AND EFFICIENCIES 



SmoT Pomp.* 


CAPAcrms, Gallons 
PBS MnnriBAT 


Effideney^ 
Per oent 


Sin of Pump,* 
Inches 


Capacitibs, Gallons 

PEB MiNUTB AT 


Efficiency, 




10Pt.VeL 
perSeeondt 


12Pt.VeL 
per Seeondt 


10 Ft. Vd. 
per Second t 


12 Ft. Vd. 
per Seeondt 


Percent 


1 

3 
4 


26 

66 

98 

220 

892 


• • • 

• • • 

• ■ • 

264 
470 


27 
86 
43 
60 
66 


6 

6 

8 

10 

12 


612 

881 

1,667 

2,448 

8,626 


734 
1,068 
1,880 
2,938 
4,230 


59 
62 
66 
67 
69 



^Alao liae of diacfaarfe outlet and antUett diameter suction inlet, 
t Vdoeity throuKh dlscharce outlet. 



40 80 




600 900 1200 

Capacity In Gallons per ^Ingta 

FlQ. 15. 



1800 



JVohite Pomp. The ordinary type of centrifugal pump, Fig. 16| with a single impeller and 
without guide vanes in the casing, the latter having the shape of a volute curve, is known as 
a volute pump. 

The closed impeller volute pump is usually provided with a double inlet which removes any 
end thrust that results when only a single side inlet is used. 

This type of pump is used for heads ordinarily not exceeding 65 ft. to 100 ft., depending on 
sise of pump, as beyond this point the efficiency rapidly falls off. 

Turbine or Multi-Stage Pumps (Fig. 17). For higher heads or pressures, the impeller or 
runner is of the enclosed type and guide or diffusion vanes are introduced in the casing in order to 
direct the flow from the runner and increase or raise the efficiency by transforming a larger propor- 
tion of the energy, which exists in the kinetic form at the outlet of the impeller to the pressure form 



POWER PLANTS AND REFRIGERATION 




TrPK Pump. 



END SECTIONAL ViEW 



SIDE SECTIONAL VIEW 



Pio. 17. TVRBIKB TtPB Pump — SnraLE Stags. 



PUMPS 



363 



and reduce Uie loea of head in Uie pump casing to a minimum. A single runner is now used for 
total heads up to and including 150 ft. or 65 lb. per sq. in. pressure. The efficiency of this type, 
siiigle stage, varies from 50 to 70% when operated at the rated capacity and head. 

For pressures above 50 lb. per sq. in. the pump is constructed with two or more runners 
or stages depending upon the pressure. Approximately 50 lb. per sq. in., or 125 ft. head, for 



Multi-stage pimips direct connected to steam turbines or motors are now largely employed 





Pig. 18. PARABOLoro of Revolution. 



Fro. 19. Elementary Centrifuqal Pump 
WITH Throttle Valve Closed. 



for boiler feeding and similar high-pressure service. The general arrangement, operation data, 
and overall dimensions are given by Fig. 32 and Table 21. 

General Theory of Centrifugal Pumps. If the cylindrical vessel, Fig. 18, be filled with a 
liquid to a level A A and then set in rapid rotation on its axis XX t the liquid will finally assume 
the same angular velocity as the vessel. Consider the small particle oT liquid M, whose weight 
is say W. It is at a distance r from the axis and will be subject to a centrifugal force C = W a* r/g^ 
where a is the angular velocity and W the force due to its own weight. The resultant will be R, 
If the surface <^ the liquid at Af is to remain in equilibrium under these forces the tangent to 
the surface at M , 77, must be at right angles to the resultant R, 

The equation of this surface nuiy be established as follows: The point P is at a dis- 
tance X from the axis and at a distance y from an axis, which can, for convenience, be taken 
through the lowest point of the surface at'^. If the ratio between y and x is known over 
the whole range — that is, from the axis to the walls of the vessel — the equation of the curve 
is determined. Now from the triangle of forces 



dy 
dx 



Centrifugal Force Wa* x aH 

9 



W 



gW 



dy — — xdx 
9 



and 



1/ = / — xdx = — x 

^ Jo g 2g 



That is, the resulting surface will be a paraboloid of revolution. 

The head on the base under a point, as N, will evidently be greater by A« than the head 
h^ under B. This head is due solely to the peripheral velocity, um &t iV, because the position 



364 POWER PLANTS AND REFRIGERATION 

of the particle N is due to the energy of rotation. The potential energy imparted to the particle 
is Wh^ and this must be equal to the head corresponding to the peripheral velocity of um* 

Suppose the open vessel just considered is converted into a closed vessel (Fig. 19). A 
paddle-wheel for rotating the liquid at a uniform velocity is provided. Its object is simply to 
rotate the liquid at a uniform velocity, not to displace it. Suppose when the liquid is quiet it 
stands at the level A A in the manometers. Evidently the head on the whole base will oorreqxMid 
to the difference between the levels A A and FF as it did in the previous case. 

Now let the paddle-wheel be set in rotation so that the velocity of rotation will be exactly 
the same as it was in the case of the open vessel. If the piezometer at iV^ is at the same distance 
from the axis as the point N in the open vessel and if in both cases the linear velocity is tiM the 
head in the piezometer will be as before ^ + h^- The head at B will be h^ In fact, there win 
be a tendency toward a formation of the same paraboloid of revolution as in the open vessel, 
so that if another piezometer be inserted at S^ the liquid will rise in it up to the point where the 
paraboloid of revolution (shown in dotted lines) crosses the line of the piezometer. 

between these two points is . Note that there has been no flow of liquid between 

the vanes, the water has simply been rotating. The case is analogous to a centrifugal pvanp with 

the discharge valve closed, and, theoretically, the expression for the increase in head could be 
used to determine the shutK>ff pressure of a centrifugal pump if the inlet and outlet diameten 
<^ the vanes and the speed of rotation were known. 

Fig. 16 is a diagrammatic section of a single stage double suction pump, with a harisootal 
instead of a vertical shaft, showing the customary arrangement of vanes, etc. Consider the two 
points 1 and 2. If the throttle valve at /C is closed and if we denote the peripheral velocities at 
1 and 2 by Ui and us respectively, a change of head, or change of energy per Ib^ equal to 

— will be obtained. 

If the throttle valve is opened this difference of head will cause a flow. The water enten 
the runner from either side, at aOf passes through the channels between the vanes or blades and 
enters the volute through the annular passage or whirlpool chamber 66. It leaves the volute 
or casing at K. Consider the points 1 and 2 on the runner (Fig. 20). Suppose the water in 
entering at ** a " has a velocity ci, and that it still has that velocity as it enters 1iie vanes. The 
direction of flow will be radial, and in order that there shall be no shock at entry, to the vanes 
the components of the radial velocity ui and uh must be of such size that Ci will be radial in the 
theoretical case. 

ActuaUy the water as it enters through the suction inlet will be in contact with the rotating 
shaft and hub and will start to whirl. The water will receive a velocity ci rotation increasing 
with the distance from the axis. The effect will be to swerve Ci slightly from the radial direc- 
tion as shown in Fig. 20. For the present this irregularity will be neglected. 

At 2 the above reasoning is reversed. The circumferential velocity is th and the 
relative velocity wt. These have as their resultant Ct, which represents the absolute velocity with 
which the water leaves the runner. 

It has been shown how the head produced by a pump when there is no flow may be cal- 
culated theoretically. When there is a flow additional terms must be considered. The law of 
conservation of energy will be applied first to the flow inside the runner channels themsdves, 
disregarding for the moment what changes of energy there may be outside the runner. 

The energy in one lb. of water at 1 is 



PUMPS 



365 



where ^ is the statie or bursting pressure at that point. During the passage of the water through 
the wheid, evidently, from what has akeady been said, there has beea added to each lb. of water 

tif' — * til' 

the energy — . Besides, there has, of course, been friction representing an energy loss 

which can be represented by a head hr. 

At 2 every lb. of water contains the energy 




Yefocity Diagrams at Inlet and Outlat 

Fio. 20. 



From the above then 



^i-h 



'2g 






Trani^poee the equation to read 



Wj« - Ui* Wi* -w^ 
Wj — /ll = 1 ;; Ay 



(1) 



(2) 



2g 2g 

and an equation tor the change in static bead or bursting pressure is the result. 

The change in edbrgy due to the fact that the absolute velocity on leaving b different than 
the abaohite velocity on entering the runner is still to be accounted for. 

The absolute otergy in a lb. of water at 1 is: 



*i + 



2g 






and &t 2 it is: 

The change of absolute energy per lb. is therefore: 

(»■ + ^) - ('■ ^ f ) - '-^ ^ *- '■' 

The total change of energy or energy added to a lb. of a liquid passing through the runner is: 



(3) 






2? 



But from equation (2) 



kt-ki+hr 



Uf — Ml' . tTi* — tC^ 



2(? 



+ 



2ff 



(4) 
(5) 



366 



POWER PLANTS AND REFRIGERATION 



Therefore 



Cj* — Ci* tit* — ttl' vh* — Wi? 

L = — z h — h 



2? 



2ff 



2ff 



(6) 



The energy, in ft.-Ib., put into one lb. of a liquid is at the same time the total hei^t or bead 
through which the liquid could be raised by that energy. It is therefore permissible to write 



H = 



Ct' — Ci' Ui* — Wi* tCt' — UJj* 



2g 



-h 



2g 



-h 



2g 



(7) 



This equation is far from satisfactory for practical use, for the only terms in it that can 
actually be determined as a basis for design are ui and Ut, which can be determined from the 

2X171 

relationship u = , where u is the peripheral velocity in ft. per sec., r the radius in ft. 

Ov 




Out/et Velocity Diagram 
FiO. 21. 



and n the speed of rotation in revolutions per minute. To arrive at something tangible c and 
w must be eliminated. 

Redraw the diagram at point 2 in Fig. 20 as shown in Fig. 21. Evidently 



and by analogy 



t£7j« a Ci' -h Wj* — 2t48C2 C08 ttf 



it^« » ci' -h ui* — 2wiCi co« ai 



H = 



2u2Cs co« as — 2uiCi cos a i 



(8) 



For purposes of design it is accurate enough if wc consider that angle ai = 90°, that is, that 
the water enters the runner radially. Then cos ai ^ O and 



// = 



tijCj cos a 2 
9 



(9) 



One of the things usually known to the designer, or at least assumed by him, is the outlet 
angle of the runner j9s. It can be introduced into our equation as follows: 



Ct 



sinfii 



sin Pi 



Ct COS at 



ut sin [180** — (ax -h fit)] sin ai cos fit -h cos aj sin /8j 
Ut cos as sin fit i^ 



sin at cos fit + cos as sin fit Uin as cot fit +1 



u-,^ ( 
Substituting in equation (9) H ^ -^ \ 

Q \ 



g \tan at cot fit -\- 



T.) 



(10) 



(11) 



PUMPS 



367 



The theoretical velocity required to produce a certain head is 



Ut ^ yg H (tan at cot fit + I 



(12) 



In this equation Ut and cot fit are fixed quantities, but tan at is a function of the quantity of water 
delivered by the pump. For if the quantity of water delivered by the pump changes, cv will 
change in proportion. Ck>nsequently at will vary. To simplify matters, the designer writes the 
equation 



Ut = Ku^2gH; K, 



Ut 



1 



'^2gH 



H = 



Ut' 



2gK, 



(13) 



and detennines the values of K^ over the whole range of the pump at a given speed by actual 
test. By proper choice of the coefficient K^, which has been calculated from a test of a similar 
pump, di^neter and speed, at which another runner of similar type must be run for any desired 
head, may be determined. 

The quantity of water delivered by a pump is known as soon as <v ai^d the circumferential 
outlet area of the runner are known. The value of (V is usually determined from data taken 
on the test floor. As the quantity of water delivered varies, the head on the pump varies and the 
foUowing equation may be written. 



Cr^K^ ^2gU', K„^ 



Cr _ 

^2gU' 



2gK^* 



(14) 



and the values ci K^. over the whole range of the piunp under test can be determined and kept 
in the form of curves as shown by Fig. 22. 

It is frequently necessary to calculate what a pump will do when run at another speed than 
the one at which the test was made. From equation (13) it will appear that the head varies 






070 
0,60 

"^7^0.40 

^ 0.20 





y 


— 


^ 


■-. 


>! 












i 


f 








^ 


v' 


^ 


^ 


>= 


^ 


/ 
















^ 


7 








i^ 


^ 


^ 














^ 


^ 


















/" 




















t 






















ti. 


1 1. 


i 1. 


3t. 


4 h 


i\ 


.BL 


7i. 


s/. 


Ti 


1 



^0.18 



tm 9 w^ 99V w9~w- rm\^ ««ir *>•# ««^ v'w «• 

K^ {Coeffichnt Jmpeller PariphajutJ VeJ^ 



5 a/7 



'no.12 

^ 0.1 1 



0.16 
0.15 
0.14 
0.13 



































^ 


k. 
































\ 


^ 
































\ 


\^ 
































„.^ 


N 


































^ 


■^ 


^ 






































— 






— . 




_ 


/ : 


> i 


4 




; £ 


' ; 


' I 


1 s 


> /( 


/ 


/ A 


2 1 


3 1 


4 1 


5 1 


8 



Fig. 22. 



$i^6 of Pump to lacbet 
FlO. 23. 



as the square of the speed. The quantity delivered varies as (V and c^ varies with the speed Uf 
Both di these statements hold, however, only within reasonable limits. The power input varies 
as the cube of the speed. 

The above theory holds for any type of single-stage pump. Multi-stage pumps are simply 
single-stage pumps in series, although the constants used in designing them are different than 
those used with single-stage pumps. The head to be obtained from two runners in series is 
twice the head to be obtained from one alone. The quantity of water delivered is, however, 
that to be expected from one runner. 

The following matter, referring to Figs. 22 to 25 inclusive, was taken from the article " A 
Rating Chart for Centrifugal Pumps,'' by L. 0, Bradford, appearing in the ^'Engineering News," 
VoL 72, No. 8. 

When a new line of pumps is to be designed, a pump having the desired shape of character- 
istic is tested, and the " K^ — Ker " characteristic is plotted as shown in Fig. 22. Next, the 



368 



POWER PLANTS AND REFRIGERATION 



values of the coefficient of discharge velocity for the various sizes of pumps to be made 
Bumed. The values of these discharge coefficients, 



■42gH 



(1) 



where V4 » velocity of water in the discharge pipe in feet per second, are such as former experi- 
ence has shown to give good results. Fig. 23 shows graphically how the values of the discharge 
coefficient may vary with the size of the pump. EUtving assumed the discharge coefficient, the 




/ 2 3 4 5 6 7 8 9 10 H 12 13 14 iS 19 

8lzB In hch«9 

FlO. 24. 



capacity of a pump for a unit head can be computed. The following notation will be used: 

Q = pump capacity, cubic feet per second; 
iV = r. p. m.; 

K4 = discharge coefficient = T/^^ 

d = diameter of discharge, inches; 
H = head per stage; 

aie 




40 



60 63 7Q,5 100 128 159 200 251 316 400 500 630 795 iOHO 

qss {Gallon per Minute) 

FlO. 26. 



A B area of discharge, square feet, = 



0.7854 d* 



144 ' 

gal, per mm. 
q = capacity of pump at unit head — ~7^ 



Now 



PUMPS 369 

- , . , r. p. tn, 
n = speed of pump at unit head = — . ; 

Di B outside diameter of vanes, inches; 

Wt » width of impeller passage at exit, inches; 

T « t3rpe of impeller = =-; (2) 

Wt 

WgX231 

^ 1728 X 60 ^ ' 

also V4 ■* -7, and by substitution, 
A 

^ q VFx 231 X 144 
** " 1728 X 60 X 0.7854 (/« 

Now, by dividing both sides of this equation by '>l2gHf simplifying and remembering that 
K4 = , we get 

/^rf =j^,org = 19.6A'rfrf« (6) 

From this equation the value of the discharge at unit head, or, as it is more frequently called, 
the " unit capacity," may be computed as soon as values are assigned to K^ and d. The speed 
in feet per second at unit head, or " unit speed," is found as follows: 

K^ ^ ^^«^ . ^ ^ KuXl2 X V2^X60 

60X12X V2^£f *' " ' 3.14 X/>t 



Dividing by V H and reducing, 



Ku X 1838 



Were the values of K^ and Dt known, the chart could now be laid out. But Dt is usually 
unknown unless the chart is being drawn for an existing line of pumps. Of course Dt could be 
assumed, but it is usually preferable to assume the type T of the impeUer. 

Assuming that 85% of the total area of the periphery of the impeller is available for the 
passage of water, we have 

gal. per min. X 231 

' "'Cr X 12 X 60 X ^ I>i X 0.85 

Dividing the numerator and denominator on the right by V ^ 

0.015 a 

^• = ^ (^> 






Substitutmg -7 for Wt ^ ^ \i ^ (8) 



Placing this value for Dt in equation (6) there results, 



w = 7-^ (9) 

<qT 



370 POWER PLANTS AND REFRIGERATION 

Values for q and n may be obtained for any size of pump using any type ci impeller, the 
values of Ku and Kcr being taken from the test on which the line is based. Four values of q 
and n are usually computed for each impeller represented. The points usually taken are the 
point of maximum eflSciency, 90% of maximum eflSciency — at a capacity lower than that at 
which maximum efficiency is reached, and 95% and 90% of maximum efficiency — at greater 
capacities than for maximum efficiency. 

Example. The use of these formulas can best be made clear by sample computations. Take, 
for example, the case of a 6-in. pump, 
d = 6 in.; 
Kd — 0.132 (based on tests of existing pumps which give good efficiency. Fig. 23); 

7 » 23 (assumed); 
Ku = 1-25 (from test, Fig. 22); 
Kcr = 0.22 (from test. Fig. 22); 

Calculation of q and n for 100% of Maximum Efficiency. 
Maximum Efficiency = 67% (from test, see Fig. 22). 

q = 19.6 X 0.132 X 6 X 6 = 93 

15,016 X 1.25 X 0.47 

n == 777. = 192 

46.2 

Calculation of Other Than the Point of Maximum Efficiency. Since Ku and K„ are re- 
spectively directly proportional to the peripheral velocity of the impeller and the radial velocity 
of the water, it follows that n and q are, for a given pump, directly proportional to Ku and Ken 
respectively. If, therefore, it is desired to find the values of q and n with the pump operating 
at other than maximmn efficiency, it is necessary only to write the following equations and solve. 

nx =* — -^ UO) 

■ 

«' = -rT^ ^"^ 



If it is desired to find the values of q and n when the pump is operating at 90% of its maxi- 
mum efficiency, and delivering less water than it would at maximum efficiency, 

efficiency = 0.90 X 0.67 = 0.603 

By reference to the impeller-velocity characteristic (Fig. 22), it is seen that the values of Ku 
and Kcr corresponding to an efficiency of 60.3% and a minimum discharge are Ku = 1.15 and 
Kcr =■ 0.155. Substituting in equations (10) and (11), 



192 X 1.15 93 X 0.155 

1.25 " ^^ ^^" 0.22 



n^ = ~-~^- = 177 qgo = — ~^ — - 65.5 



The points on the front of the curve, that i& when the pump is discharging more than its normal 
amount of water, are found in a similar manner. Fig. 25 shows such a chart plotted for a com- 
plete line of pimips, sizes varying from 5 to 16 in., and for two tjrpes of impeller for each size. 
C7«e of Chart. When an order for a pump is received the values of q and n are inunediately 

gal. per min. 
calculated from the conditions of operation, by means df the equations q — 7= and 

n = -^~=-^f and the point plotted. The impeller having the next value <rf njbelow that found 

is taken, and if the difference is large the tips of the vanes are cut back. (This is done to bring 
the speed up to that required. It is readily seen that cutting back the vanes increases the speed 



PUMPS 371 

when it is lememberad that for any ^ven impeller the outside diameter of the vanes must at- 
tain a certain peripheral velocity in order to produce a certain head.) 

Suppoee, fc»- example, that an order came in for a pump to deliver 1500 gal. per mtn. sgainat 
■ bead of 250 ft. when running at a speed of 1760 r.p.m. The pump will obviously consist of 
two stages, for while a head of 250 it. per stage is possible, it can usually be attained only at the 
surifioo of efficiency. The head per stage is then 125 ft. 

1500 „ 1760 

q --;= - 135 n r-^^ = 157 

Vi25 Vl25 

The point determined by these values of g and n is then plotted as shown on Fig. 25. The 
impeller choaea will be the one having the next highest values of q and n. In this case an 8-in. 
pump having a type 23 impeUer would be chosen and the vanes cut back a trifle. The pump 



FlO. 2e, AaaAKQBHENT OF APPARITCS FOR CENTRIrUQAL PUMF TEST. 

would operate at about 98 per cent of its maximum efficiency. From Fig. 24, it is seen that the 
maximum efficiency of an 8-in. pump operating at 125 ft. per stags is 71 per cent. The eflScicncy 
ot the pump selected would therefore be about 71 per cent X 0.98 — 69.5 per cent. 

Teatiiig and Rating Centrifugal Pumps.* In testing centrifugal pumps the main measure- 
ments usually made are: 

1. Power input. 

2. Quantity of water pumped. 

3. Total head. 

4. Speed. 

Power Input. The pump niay be either direct-connected to an electric motor or driven by 
belt from a transmission dynamometer which in turn is driven by an electric motor. 

In the first case, if it is a direct-current machine, the motor will have to be tested at the 
various speeds at which it is to be run and at each speed a curve is to be drawn showing the ratio 
between the brake horsepower developed and the watt input into the armature over the whole 
range of power which the motor may have to develop at that particular speed. It is well to 
keep the field current at some constant value over the whole range for any given speed and to 
take care that when the motor is operated at this speed the same field current is maintained. 
Knowing the electrical input into the armature, it is an easy matter then to choose the corre- 
sponding brake horsepower from the curve. 

* The muthon unt indebted to Edwin Fnuik lar the fallowing initter. 



372 POWER PLANTS AND BEFRIGERATION 

The Lewis type dynamometer sSatda another means <^ meaBoring the power input into 
the pump. It is ^own at D in Fig. 26. 

Power is transmitted from the motor M to the fixed shaft L. On this shaft is mounted 
the gear R which communicates its motion to the gear wheel R', mounted on the shaft L^ The 
latter rotates in a lever or cradle Q hinged to the main f ramft by means of a steel plate spring 
at A. The shaft L' transmits its motion through two Hooke's joints to the fixed shaft on which 
the pulley is mounted. From this pulley the power is transmitted by belt to the piunp. 

Let p = the effective belt pull, 
r B radius (^ the pulley Q. 

When power is transmitted the effective U»tiue (n* moment is pr. This tcnrque is equal to 
the moment p'f. The following relations can be written: 

f <* + & „ 

V - -— - P 



pr«p"(^X/) 



p" »~Tr 

c 



TT ( — X — — X— I = IT X constant. 
\ c o r/ 



The wei^t W can be moved in and out on the lever d which is provided with a scale gradu- 
ated to read directly the pounds of effective pull on the pulley Q, Theoretically, then, the hone- 
power input into the pump is 

p X2yr Xn 

33000 

There is a small loss in the belt which should not exceed 5 per cent of the power trans- 
mitted. 

The quantity of water pumped may be determined either by the V-notch weir method, 
or by means of a Venturi meter. 

The total head will be determined as described in the general notes on pump tests. In- 
stead of a suction gage, a mercury manometer is generally used, as it is far more sensitive. Every 
reading taken must be corrected to give the pressure at the center of the pump. This means 
that if the manometer is below the center line of the pump, a head of water equal to the vertical 
distance between the horizontal center line of the pump and the top of the mercury on the side 
of the manometer, subject to the vacuum, must be added to the head in ft. of water corresponding 
to the difference in level of the mercury. If the manometer is above subtract that head from 
the indicated suction head. 

The revolutions per minute of the pump may be measured by both tachometer and speed 
counter — one to be used quickly to adjust the speed of the pump, the other to serve as a check. 

General Method of Condvuiing Test, Fig. 26 shows the arrangement of the testing apparatus 
for a test in which the Lewis dynamometer and the Venturi meter are used. M is the motor trans- 
mitting power by belt to the djmamometer D, from where it is transmitted once more by belt 
to the pump C. The pump draws in water from the sump tank and delivers it past the throttle 
valve T into the discharge main. After passing through the Venturi meter V the water is returned 
to the sump tank. 

The water in the S3rstem will have a tendency to heat because the various heads on the 



PUMPS 373 

pomp are produced by throttling with the valve T, To get a maximum amount of water in 
the system Uie tank 8 must be filled quite to the top at the beginning of the test. To help cool 
the wiU;er further a cock is provided at E^ permitting some of the hotter water to be drained off 
to the sewer and cold ¥rater to replace it can be obtained from the pipe F^ which is connected 
to the mains. The temperature of the water in the sump should be maintained constant by 
providing a steady flow at both E and F, 

Centrifugal pumps usually operate at constant speed, but the head, the quantity of water 
ddtvered, the horsepower input and, therefore, the efficiency vary widely, althou^ systematically. 
To deariy understand Uie interrelation of these variables, curves should be plotted. The data 
fcv each is taken simultaneously and at the same speed of the pump. One shows the relation 
betwe^i capacity and head, one between capacity and power input, and the third the 
relation between ciq>acity and efficiency. In aU of these the capacities are plotted along the 



Fig. 27 shows three such sets of curves from an 8" single-stage pump. The " head " curve 
shows the variation of total head as the ci4)acity is increased from xero up to the maximum 
quantity that can be forced through the pump at the given speed, with the discharge valve wide 
open. The ** horsepower input " curve indicates the variation in the power required to keep 
the pump up to the required speed simultaneously with the changes in head and capacity. The 
efficiency curve follows from the other two. * 

From the ** head " curve at any given speed can be obtained at once the actual head against 
which the pump can deliver a given quantity ci water. But it is worth emphasizing that in 
order to give a certain quantity of water at a given speed the pump must be working against 
the head, as indicated by the ^* quantity-head " curve at that capacity. Suppose the pump is 
running at 1800 r.p.m. and it is desired to deliver 400 gallons per min. against 200 ft. head. 
From our curve to deliver 400 gal. per min. the pump must be working against 214 ft. head and 
not 200, for at 200 ft. head it would deliver 1375 ^ per min. To produce the 214 ft. head 
the excess 14 ft. must be produced by closing the throttle valve partly. The only other yny 
in which we could have obtained the desired point of operation would have been to reduce the 
q)eed of the pump until the desired head and quantity would be obtained. 

Conversely if we have such curves for any pump and operate it at the test speed we can, 
by observing the suction and discharge heads, determine very nearly what quantity of ¥rater 
is being delivered. 

In goieral, "quantity-head" curves rise more or less from the shutroff point, as at Oot for 
a part di the total range; and then after passing the maximum, fall with increasing capacity. If 
the curve can be continued far enough it will be found that the head finally comes down to zero 
along a line approaching the vertical. It is to be noted that the curves at various speeds are 
practically pandlel cnr concentric, and if, for instance, we wanted to construct a "quantity-head" 
curve fcnr a speed of, say, 2000 r.p.m. we could transfer any point on the 1800 r.p.m. to the 
new curve by remembering that the heads vary as the square of the speed, and the quantity 
dehvered directly as the speed. 

The " horsepower input" curves will be found below the " head " ciurve. At 1800 r.p.m. it 
will be found that 800 gal. per min. will be delivered against 212 ft. head. To find what horse- 
power is required follow the 800 gallon line until it intersects the 1800 r.p.m. " hp. input" curve. 
We find that 83 hp. are required. This would not, however, justify us in purchasing, say, an 85 
hp. motor for the work. For if the discharge line should burst the head would gradually fall 
off along the "quantity-head" curve. When the head has dropped to, say, the point O4, the horse- 
power to be developed by the motor is 119 hp., as indicated by the point P4. A motor capable of 
giving that horsepower would have to be selected in oiu* case. In general the " horsepower" 
curves rise to a maximum and then generally fall off abruptly to the same hp. input as is required 
at shut-off. 

The efficiency curves lie, in this case, between the head and the power curves. They start 
fnnn zoo, at zero capacity or shutK>ff, at which point the pump does no useful work, althou^ 



374 



POWER PLANTS AND REFRIGERATION 



it consumes power which is entirely wasted in friction. As the capacity increases the efficiency 
increases until it reaches a maximum, and then decreases to zero at the full capacity of the pump, 
where again no useful work is performed as the head against which the water is pumped is xero. 
In general, the curve is a semi-ellipse, with the capacity line as the minor axis. It is worth noting 
that up to a certain limit the maximum efficiencies increase with increasing ^)eed, only to decrease 
again when that maximum has been passed. The point of maximum efficiency moves to the 
right or in the direction of increased capacity as the speed is increased. 




20 ^^ 



QUANTITY-HEAD. EFFICIENCY AND H.P. INPUT 
CURVES FOR 8"SINGLE STAGE CENTRIFUGAL PUMP. 



1 



t 

I 

« 




400 



8U0 



1200 1600 

Gal. per ml a. 

FlO. 27. 



2000 



2400 



2800 



The data for a set of such curves are to be taken as follows : The pump is primed and brought 
up to the predetermined speed. As soon as there is a decided flow the throttle valve is closed. 
In this condition the total head is determined and also the hp. input. This locates the p<Hnt 
Oo on the " quantity-head " curve and also the point P© on the " power " curve. The efficiency 
is, of course, zero. 

The throttle valve is now opened wide and the speed adjusted by regulating the motor. 
The total head on the pump will now consist of simply the suction head and the few feet of head 
necessary to discharge the water through the pump, the open valve and the discharge piping. 
Under these conditions the maximum possible quantity of water will be delivered. In this way 
the point Oi on the " quantity-head " curve and the point Px on the " power " curve will be 
located. The point ^i on the " efficiency " curve can be located by calculaticm. 

We shall now have to explore the region between the points Oi and Oo. A good way 
is to find the discharge gage reading at Oo and at Oi and divide the range roughly into. 



PUMPS 



375 



aay, eight equal partfi. Now gradually close the throttle valve until the discharge gage reads 
the first of the pressures decided upon. Regulate the speed and take the readings. Then pro- 
ceed to the next point. It is essential that at least the '' quantity-head " curve be plotted as 
the test proceeds, otherwise there may be gaps that will be hard to fill in when the curves are 
being drawn up in the computation room. 

When the last of these points — ^for example Os, and Pt in Fig. 27 — ^has been determined, 
change the speed of the pump to the second predetermined speed and proceed in the same way. 



200 



160 



^20 






Z80 



40 




QUANriTY-HEAD. ISO-EFFICIENCY 
^ND BRAKE HORSE POWER CURViS 

OF AN 

8" SINGLE STAGE CENTRIFUGAL PUMP. 



400 



800 



1200 leoo 

Gal* per mla. 
Fig. 28. 



2000 



2400 



2800 



The result will be another series of similar curves, as shown in dotted lines. By running at 
still another speed, ciurves, shown in dot and dash lines, will be obtained. 

When the above results have been plotted the so-called '^oak tree'' curves (Fig. 28) can 
be constructed from them. 

In Fig. 27 the 70% efficiency Une crosses the efficiency curve of the 1800 r.p.m. test in 
the points ciei. From ciei draw verticals to the " quantity-head " curve. The intersections are 
mimi. Now find where the 70% efficiency line intersects the two other curves. The points are 
eies and e^t. On the corresponding *^quantity-head " ciurves the corresponding points are mtmt 
and nnnit. 

In Fig. 28 the " quantity-head '' ciurves are reproduced and smooth curves are drawn through 
the points mimi, r^imi, and niimi. Id an entirely similar manner curves are drawn for 65 and 
60% efficiency. 

In Fig. 27 the 65 horsepower line intersects the power curve of the first test at rin. Pro- 
jecting to the corresponding '' quantity-head " curve the points nitii are located and those in 



376 



POWER PLANTS AND REFRIGERATION 



turn are transferred to Fig. 28. On the second power curve the points r^t and on the third the 
points fifi were located. The corresponding points on the *^ quantity-head ^* curves are n^n* 
and nsiti. Transferring the points nitti, nifts, and nin>% to Fig. 28, and drawing a dotted line 
through them all we have located the locus of all points at which it requires 65 hp. to keep 
the pump up to the required speed. In a similar way the other h(»sepower curves are located. 

Example. Curves like these for each type and sixe of pump are issued to the branch sales offices 
of the larger centrifugal pump concerns and are used as follows: Suppose a customer wants a pump 
to deliver 1800 gallons per minute against a head of 140 ft. The salesman sees at once that if the 
pump is driven at 1650 r.p.m. it will give the desired quantity quite readily. He knows, that 
he can guarantee 70 per cent efficiency and that about 83 d.hp. will be required at the given load. 
To find the sixe of motor required for the maximum possible load in case of accident to the pipe line 
he must consult Fig. 27. There he will find that the maximum load that can come on the motor will 
be 91 hp. Consequently he will probably choose a 90 or even a 100 hp. motor in order to be on 
the safe side. 

Steam Consumption of Steam Turbine Driven Tuil>ine Pumps. Table 14 contains the 
results of tests on the comparative steam consumption of turbine driven and reciprocating boiler 
feed pumps on three vessels of the U. S. Navy, where the best pumps of each type were in use. 



TABLE 14 



Seout Cruiser 


Steam 

PraiB, 

Pounds 


Back 

Press, 

Pounds 


G.P.M. 


Head 

in 
Feet 


Watbb Ratb 

PER W. Hp. 


AdTsntaffs 

in Favor 
of Turbines 


Redp. 
Pumps 


Turbine 
Pumps 


Birmingham, 

Salem 


188 
203 
187 


6 
6 
6.7 


329 
193 
219 


488 
690 
610 


' 88.0 

91.8 

101.0 


61.8 
61.2 
64.0 


26.6% 
33.6 


Ch€8Ur 


84.6 






Average 


* 
• • • 


• • • 


• • • 


• • • 


91.9 


62.4 


81.2 



NOTES ON THE INSTALLATION OF PUMPS 

The following matter regarding the installation of pumps is an extract from a bulletin issued 
by the Cfotdds Mfg. Co. 

Reciprocating Pumps. Locaiion, The pump should be locate^l as near the source of supply 
as possible. Never under any circumstances exceed the height and distance from the source 
of supply, given in Table 1. When hot and thick liquids are to be handled, they should always 
flow to the pump. It is always advisable to place the pump so that it can be readily reached 
from all sides. 

Foundaiion, Power pumps are self-contained and are not dependent upcm foundations to 
maintain correct alignment of the working parts, but a good foundation is essential to preserve 
the correct position of the pump in relation to its driving mechanism, and to avoid imdue strain 
upon the pipe connections. The pump should, therefore, be placed upon a level and secure foun- 
dation. For many smaller pumps, a good plank floor, such as is found in any well-built mill or 
factory, is sufficient, but for the large, heavier pumps, a substantial concrete foundation with 
anchor bolts should be provided. Where such foundations are to be made, plans will be sent 
so that the foundation work can be completed before the arrival of the pump. 

Piping in Oeneral. Rim all piping in as direct a line as possible; avoid all unnecessary 
turns. See that all joints and connections are tight, and if the pipe lines are long, use larger sixe 
pipe than that listed for the pump, in order to reduce friction and to keep down the pressure on 




Fig. 29. Sinolb-Staob Double-Suctiok Oodld Puiir. 



TABLE 15 
DIMENSIONS OF GOULD'S SINGUl^TAGE DOUBLE-SECTION PUMP. 



the pump. This will prove economical in the long run. To illustrate: Suppose & pump is dis- 
chargiiig 100 gsUona per minute through & 3-inch pipe, 1000 feet in length. From the frictiiMi 
table Tor iron pipe, it will be seen that the toss <A bead for 100 feet of 3-incb pipe discharging 
100 G. P. M. is 3.52 feet, which for 1000 feet of pipe means a loaa of 35.2 feet. U in pUce of 
24nch ppe, 4-inch pipe were used, the corresponding loss would be 0.S8 feet per 100 feet of length, 
or 8.8 feet for 1000 feet of pipe. The saving in preeaure on the pump would, tiierefcwe, be 35.2 — 
8.8 ^ 26.4 feet, which is equivalent to approximately 11.4 lb. preesure per sq. in. 

Suiium Pipe. The suction pipe should in no case be smaller than the siie given in the 
pump table, and if very long, it should always be larger. In laying the mictioa pipe, a unif<nm 
grade should be maintained throughout to avoid air pockets, and if possible the lines should 
have a drop of not less than 6 inches in each 100 feet length toward the source of supply. 



378 



POWER PLANTS AND REFRIGERATION 



Vtieuum Chamber. The addition of a vacuum chamber greatly aids and steadies the suctkm 
flow, «nd one should be used on high suction lifts. Where it is possible, the vacuum chamber 
should be mounted on the pump, on the side c^posite the suction intake. 

Foot Vakfe. When the total suction lift exceeds 15 feet, or the suction line is over 100 feet 
in length, it is advisable to place a foot valve on the end of the suction pipe. This keeps the 
pipe and pump chamber filled with water, thus avoiding the possible necessity of priming the 
pump each time it is started. 

Strainer. If the source of supply is at a point where foreign substances may be drawn into 
the pump with consequent clogging of the valves, a strainer of good liberal area should be used 
on the suction pipe; preferably one that can be examined and cleaned occasionally. 



TABLE 16 

SPEED TABLE FOR GOULDS SINGLE-STAGE, DOUBLE-SUCTION CENTRIFUGAL PUMPS 





Speed 
Max. 








R. P. M. POB Total Hbaos fbom 10 to 


160 For 


Siie 


«Bd 


































Min. 


10 


20 


80 


40 


60 


60 


70 


80 


90 


100 


110 


120 


180 


140 


150 


2S. . . 




Max. 


1600 


2260 


2770 


8200 


8580 


8980 


4240 


4580 


4800 


6060 


6810 


5650 


5780 


6000 


6S00 






Min. 


800 


1180 


1885 


1600 


1790 


1960 


2120 


2260 


2400 


2580 


2660 


2780 


2890 


8000 


8100 


2L... 




Max. 


800 


1180 


1885 


1600 


1790 


1960 


2120 


2260 


2400 


2580 


2660 


2780 


2890 


8000 


SlOO 






[Min. 


490 


695 


855 


985 


1100 


1210 


1800 


1890 


1480 


1660 


1685 


1710 


1775 


1845 


1910 


8S... 




Max. 


1140 


1610 


1970 


2280 


2540 


2800 


8020 


8220 


8420 


8600 


8780 


8960 


4110 


4260 


4410 






Bftin. 


710 


1005 


1230 


1425 


1590 


1745 


1880 


2010 


2140 


2250 


2360 


2470 


2670 


2660 


2760 


8L... 




Max. 


710 


1006 


1280 


1425 


1590 


1745 


1880 


2010 


2140 


2260 


2860 


2470 


2670 


2660 


2760 






IMln. 


425 


605 


740 


856 


955 


1046 


1130 


1210 


1280 


1850 


1416 


1480 


1540 


1600 


1665 


4S... 




Max. 


880 


1260 


1580 


1770 


1970 


2170 


2340 


2600 


2650 


2790 


2930 


8060 


8190 


8800 


8420 






iMin. 


610 


860 


1055 


1220 


1360 


1495 


1615 


1726 


1830 


1930 


2020 


2110 


2200 


2280 


2860 


4L... 




Max. 


610 


860 


1055 


1220 


1860 


1495 


1615 


1725 


1830 


1980 


2020 


2110 


2200 


2280 


2860 






IMin. 


875 


536 


650 


755 


840 


925 


996 


1065 


1180 


1190 


1250 


1805 


1860 


1410 


1460 


6S... 




Max. 


880 


1250 


1580 


1770 


1970 


2170 


2340 


2500 


2650 


2790 


2930 


8060 


8190 


8800 


8420 






Min. 


535 


755 


925 


1070 


1190 


1310 


1410 


1510 


1600 


1690 


1770 


1850 


1925 


2000 


2070 


6L... 




Max. 


585 


755 


925 


1070 


1190 


1810 


1410 


1510 


1600 


1690 


1770 


1860 


1926 


2000 


2070 






IMin. 


846 


490 


600 


690 


775 


850 


916 


980 


1040 


1096 


1150 


1200 


1250 


1296 


1840 


6S. .. 




Max. 


776 


1100 


1845 


1550 


1780 


1900 


2060 


2200 


2330 


2460 


2570 


2690 


2800 


2900 


8010 






IMin. 


475 


670 


820 


950 


1060 


1166 


1265 


1340 


1420 


1600 


1576 


1640 


1710 


1776 


1840 


6L... 




Max. 


475 


670 


820 


950 


1060 


1165 


1256 


1340 


1420 


1500 


1675 


1640 


1710 


1776 


1840 






IMin. 


810 


485 


585 


615 


690 


755 


815 


870 


925 


976 


1025 


1070 


1110 


1166 


1196 


8S... 




[Max. 


620 


740 


905 


1045 


1170 


1280 


1885 


1480 


1570 


1650 


1786 


1810 


1886 


1966 


2080 






1 Min. 


855 


505 


615 


710 


795 


876 


940 


1005 


1065 


1126 


1180 


1235 


1286 


1880 


1880 


8L... 




Max. 


855 


505 


616 


710 


795 


875 


940 


1005 


1065 


1126 


1180 


1285 


1285 


1880 


1880 






Min. 


265 


880 


460 


535 


596 


655 


705 


755 


800 


845 


885 


925 


965 


1000 


1086 


lOS... 




Max. 


425 


605 


740 


855 


955 


1045 


1180 


1210 


1280 


1850 


1415 


1480 


1540 


1600 


1665 






[Min. 


820 


455 


655 


640 


715 


785 


850 


905 


960 


1010 


1060 


1110 


1165 


1200 


1240 


lOL. . . 




Max. 


820 


455 


555 


640 


715 


785 


850 


905 


960 


1010 


1060 


1110 


1165 


1200 


1240 




< 


IMin. 


235 


885 


410 


475 


530 


580 


630 


670 


710 


750 


785 


820 


856 


890 


920 


12S... 




fMAX. 


855 


605 


615 


710 


795 


875 


940 


1006 


1065 


1125 


1180 


1285 


1285 


1880 


1880 






iMin. 


265 


880 


460 


586 


595 


655 


705 


755 


800 


845 


885 


925 


966 


1000 


1086 


12L... 




Max. 


265 


880 


460 


585 


595 


655 


705 


755 


800 


845 


885 


925 


965 


1000 


1036 


. 


t 


[Min. 


200 


285 


845 


400 


450 


490 


530 


566 


600 


635 


665 


695 


720 


760 


775 


16S... 




fMAX. 


805 


480 


580 


610 


680 


750 


805 


860 


916 


965 


1010 


1060 


1100 


1140 


1180 




1 


iMin. 


280 


825 


896 


460 


610 


560 


605 


646 


685 


726 


760 


795 


825 


865 


885 


16L... 




^Max. 


230 


825 


895 


460 


510 


660 


605 


645 


685 


725 


760 


795 


825 


866 


886 




' 


[Min. 


170 


240 


290 


840 


880 


415 


445 


475 


505 


535 


560 


585 


610 


680 


660 



NoTB. — Speed* given Aboire Are for nuudmum, normAl, or Any intermedlAte cmtity, 
between two of the boAd TAluei given in the speed table, use the speed limits for the lower 1 



If the desired head fa 



SxpUnation of Speed Table x6. The table shows the maximuin and minimum speed of each 
siae single-stage, double-suction centrifugal pump for heads from 10 to 150 feet. The speeds given 
are for maximum, normal or any intermediate capacity. It will be seen that for each pump there are 
two sixes available. One is the high-speed pattern designated by **8" and the other the low, designated 
by **L" In every case the pattern must be selected so that the speed of the pump is within the maxi- 
mum and minimum limits. Where the speed at which the pump is to operate is not determined by 
the prime mover, we suggest that the high-speed pattern **S " be used. This is desirable because the 
higher speed of this pattern makes it possible to use a smaller diameter impeller for any head; the 
frictional iosseis are reduced and the highest efficiency is obtained. 



Water Reti^ Vaiee. A w&ter relid vkIvb of Ample sue, set at s preomm sli^tiy above that 
at wfaidi the pump ia to operate, muet be placed between the pump and anj Bhut-<^ taIvb in 
the dieofaarge [npe, in order to avoid damage in case the pump ia started with the gate valv« 
doMd. 

0<de Vaicti. Always uae gate valves, not globe valves: globe valves increase the friction, 
while gate v&Ivee offer an unrestricted passage. It ia advisable to place a gate valve at or near 
the pump ia both suction and discharge pipes, so that the valve may be closed when it is found 
iMNifWBMi^i to examine the pump. 

TABLE 17 

HOBSEPOWER TABLE FOR GOULDS SINCL&STAGE, D0UBLB-8UCTT0K CBNTfUFUQAL PUUPS 



NotH — Vvr nuiUiDd to dvtvnnlrie d.hp. ivqutnd For intcmwdlBta capvcitiea and hflsdi, asa bolow. 

KtplsniHon of HorMpower Table 17. Thia table abows the power required tor each bim dn^e- 
■tage, double-auction centrifuBal pump at both the oormal or maximum capacity acainat total beada 
from 10 to ISO feet. To detennioe the horsepower required for any intermediate capacity asaloet 
Boy desired bead, multiply the capacity desired in gallons per mioute by the total heed in feet, divide 
by the eonatBAt 4000 and then divide the result obtaimd by the efficisDcy of the pump. The rule just 
given may be ezpreased as follows: QX H 

'""'■ ■ ioooxT 

where d.hp. ia the brake honepower required, Q ii 
I deairad total head in feet, and £ ia the eflicienc 

The foUowins table of effidencies is to be used for figuring tbe horsepower required by the sin^e- 
Blage, double-euDtioii pumps. 

To allow for ample power in the driving equipmeot. low effidencies have been ^ven purpoeely. 



POWER PLANTS AND REFRIGERATION 
TABLE 18 



». 


,»^ 


Stat 


Effletauy 


a» 


EAdtDty 


Stat 




13 


1 




0.67 
0.66 

0.6S 


IS. 


0.70 




0.7Z 

si 



Fio. 30. TiMn TtmBiNE Dbivxn Bouf b Feed Pmm. 



TABLE 19 
APPROXIUATE OVER-ALL DIUBM3ION9 OF TEKSV TURBINE DRIVBN BOILER FEED PUUPS 





1 


s-rst 


R.P. M. 


,K& 




is- 




UiD. 


Uu 


E 


H 


E 


Ft 


A 


B 


c 




E 


p 


J 


i 


1;;;;;: 

e'.'.'.'.'.'. 


6 


i: 

uo 

400 


i 




2,TU 

IS 


GO 

1 


102 

8S 

1«D 


.or 
m 






t-ltf 

»-ioH 


S-«M 
s-s 




4- 

e 
e 


S 



m tor WBtUnctiiD punp*. 



N<w.4,Su 



Drain Pipe. Each cylindar is provided with suitable openings at the top to whioh small 
drAin pipes may be oonneeted for osnying irff any water that may accumulate around the Btuffing- 
bos glands. 

Cratrifugal Pumpi.* Location. Place the pump as near the source ot supply as possible and 
so that there will be the fewest possible number erf bmida or elbows in the micti<m pipe. I( possible 
the pump should be within 15 feet or less of the water level; and the suction lift diould never be 
more than 20 feet, including the pipe friction head when handling cold water. 

FountUUum. Prepare a concrete foundation with the surface ^ inch lower than the kvd 
at which the pump is to be installed, to allow for the final leveling and grouting. Have the 

* Wbtn turbioe puBun an gmployvd lor boOw rcadfac and bwidtat hot water boot hasten, tbB «BHr Amid b« 
dtUnnd to tha pmnp un<bir a htad of about T fact. 



Pig. 31. OonuM SiNaLB-AcriNO Thiplkx Plcnokb Puup. 





For 


JS 


StePump 




2Sf 


Hon.- 


SlMOlPIl). 


Oand 




:e. 


"^ 


t 


Sue. 
in 

In. 


Dto- 








7 












































ISO 


A 






B.BI 




Z8.9 
2S.9 




I 


S.etol 


BpKkllnaKfa 








































































706 


110 


B 


12 




17. «E 






10 









n 



382 



POWER PLANTS AND REFRIGERATION 



foundaUon bolts extend above the ooncrete, according to the dimenaion sheets of the pump. 
The bolts should be threaded at least 6 inches and should be set in the concrete within ga0-f>ipe 
thimbles of such size as to allow a clearance of ^ inch between the bolts and tiieir thiznbleo. 

-^ Piping in General, In selecting the piping, do not overlook pipe friction, especially if the 
pipe lines are long. All pipe friction means extra power to drive the pump, and wiiere this fric- 
ti(Mi would be a considerable item it is usually advisable to reduce it by selecting a larger m*e 
pipe, as the extra first cost of the larger diamet^ pipe will socm be returned by the saving in 
power. All unnecessary bends and elbows should idso be avoided as they increase the pipe 
friction. 




OUTLINE ELEVATION 

Fig. 32. Albsrgeb Boilbb Fbsd Pumps. 



TABLE 21 



Capacity 
BoQcr 
H.P. 



1,000 
2,000 
8.000 
4,000 
MOO 

9,000 
12,000 
15,000 



Pumps 
Nos. 



18-14-15 
22-28-24 
82-88-84 
42-^48-44 
52-53-54 
72-78-74 
92-^8-94 
122-123-124 
152-153-154 



Sin of 

Dis- 

cham 

Suction 



TET" 
2 

2H 
8 

4 
4 
5 
6 
8 
8 



Width 



Ft. In. 

80 

80 

8 8 

8 7 



8 

4 
4 
4 

4 



7 
8 
8 
8 
8 



Hdsht 



r^ 



Ft. In. 
8 1 



8 
8 

4 
4 
5 
5 



1 
10 
8 
8 





5 10 
5 10 



Lbnqth (L) fob MAxmuM Wa' 



180 Lb. 



Ft. In. 
6 
6 1 

6 11 
8 6 

8 6 

9 8 
9 6 

10 10 

10 10 



185 Lb. 



Ft. In. 

6 8 

6 7 

7 5 
9 2 
9 2 

10 

10 

11 9 
11 9 



240 Lb. 



Ft. In. 

6 6 

7 
7 9 
9 7 
9 7 

10 7 

10 11 

12 6 

12 6 



NOTB. — 100 boQcr hp. - 7.5 g.pjn. or 15% in ezcesB 
Iw len than 85% of water pranure for full capacity. 
NoTB. — ^All dinirfffMion» are approximate. 



of averafe requirements. Steam pronaure must not 



^ Suction Pipe. Never use smaller piping than the size of the pump suction opening, and 
if the suction lift exceeds 15 feet, use a larger size pipe than the size of the pump sucticm. If 
the length of the suction pipe is excessive, use suction piping at least two sizes larger than the 
suction opening of the pump, and if this is done it is advisable to use a fairly long conical reducer 
*t the pump. Never at t-pTr|pf. «. gunfmn liff^ nf ggorsttiim 20 feet u nder any circumstances. 

Always pliEtce the end of the suction pipe at least three feet below the surface of the water 
to prevent air being drawn into the pump. Avoid air pockets in the suction piping. If the 



^^"T" 



PUMPS 



383 



auetkm pipe Ib not in a vertioal podtion, it should slope downward and never upward toward 
the water, if there is any sucticm lift. 

It is desirable, especially when there is pressure on the sucticm side of the pumpi to place 
a gate vahre in the suction pipe near the pump so the capacity of the pump can be controlled 
to some extent on the suction side. It is also advisable to place a strainer on the end of the 
suctaoQ pipe to prevrait large pieces of debris entering the pump. 

^Chie and Chedc Valve. Place a gate valve and check valve in the discharge pipe as close as 
pooaible to the pump. The gate valve must be placed between the check valve and the pump. 





FiQ. 33. Albmbobb Bboulab Volxtti Pumps. 

TABLE 22 



6 
8 
10 
12 
14 
16 
18 



26 



Dit. 



5 
6 
8 
10 
12 
14 
16 
20 
24 



Max. 



700 
1,000 
1,800 
2.800 
4.000 
6.600 
7.600 
12.000 
18.000 



DmaiaiONi in Inchm 



16 
16 
18 
21 
24 
26 
28 



B 



16 H 

18 

20 

22 

24 

26 

28 

82 

84 



9 

9 

9 
11 
12 
16 
15 

IBH 
21 H 



14 H 

16 « 

18[ 

20] 

22| 

27 

81 H 



6H 

8 
11 

9 

9 
10 
11 
12 
12 



48 

48 

62! 

68 

70* 

78 ^ 

96^ 
110 



78 

98 

lllH 
124 H 
188 

141 H 

170 

190 






NoTB. — ^DIiiMBrfoiM E and G are approximato and wfll vary with liae of engine. 

The gate valve is used to control the capacity and the check valve to prevent breakage of the 
pump casing from water hanuner. This is important and necessary. 

V Electric Drive, Do not attempt to operate a motor-driven centrifugal pump from a trolley 
circuit as the line voltage of such circuits is exceedingly variable. The speed of a direct-current 
motor varies almost directly with the voltage, and the capacity of a centrifugal pump varies 
greatly with changes in the pump speed. If the pump is designed to run at the speed corre- 
sponding to the motcHT speed at maximum voltage, it will deliver little or no power when the 
voltage is low. If designed to give the desired capacity at the motcHr speed at minimum vol- 
tage, the motor wiU be seriously overloaded when the voltage rises to its highest point. 

"^ Prindng, Centrifugal pumps must be filled with water and the air removed from the casing 
(primed) before starting. Any of the following methods of priming may be used: 

!• The pump may be set bek>w the water level, in which case the water will flow throuc^ 



38ft 



POWER PLANTS AND REFRIGERATION 



the suction pipe into the pump by gravity, 
thus filling the casing and forcing out the air 
through the cocks provided. 

2. A small by-pass pipe aroimd the check 
valve in the discharge pipe may be used to 
fill the pump with water from the discharge 
pipe in situations where the discharge pipe is 
k^t full of wter. 

3. The pump may be filled from an in- 
dependent source of supply such as a tank 
placed above the pump or from a supply pipe. 

4. An air or steam ejector may be used 
to draw water up the suction main, the dis- 
charge line being closed. 




I 
S5 



I 

I 



I 



I 



Kl 












04 






I 






"3r 



■^3r 






"3!5S 

.14 .M w4 w4 r^ pt e) 



to C« A O -« M ^ V «0 -^ 

^4 V^ .M W4 .i4 .^ Ol 



M 



SSr«oooio>ocS$So» 



ao^too<HioM(0e«e<i>o 



'H^^fi 



Q0^u9«ocQ>oci^eeQ0<« 
eo^^«ot«r«aoeb«-iN^ 



eo 



tao««*o^Meo>o^eoo> 



M 



cueoioF-eotocov-iAaow 
eoeoco^2io«t>aoAO 






t0t0fc*OOflOOOOOOOOOO^ 



^ 



le to to r- 1> t> t> 00 00 e» 04 



M 



«:s? 3:;s; x:s: 



^^<«iete<0toi«t-ao^ 









04 



p-«*2::2i2Sass 



I 



ieftt-t>ooo<Hibr-eeQ 






7 



Ife 



ooooe.M^eo<«<«oooM 



1^^^3^ SST" 

t-oooe4^t«t-O0stoei 



— 3r~3I! 

e»oe4^toi«o»e3t-oto 






«-••-• e4^»o 



:«; 



eie4co<«io<DQOoe4^<D 



i 



J 

9 



8J 



^ 



000 



f 00 000 
90000 

80 000 

70 000 

60 000 
,50 000 

'40000 
'80000 



20 000 



— 18 000 



— 10 000 
Z- 9000 
— 8000 

- 7000 

- 8000 

^.^6000 
—4000 



•8»00 



'2000 



— 1600 



~^000 

:^900 

1^800 
^700 

-800 

I 600 

-480 

iOO 

1^880 

^—800 



I 

i 



/so 



100 
90 
80 

70 
80 



•98 
-90 

•84 

•78 
-72 

.88 

.60 



•64 
-48 



^2 



-38 



•80 



'24 



20 


1 




o 


18 


^ 


18 


$ 




S! 


14 


t 




!<• 




O 


12 


1 


•10 


1 




^ 



•5 



'8 

-2.5 



% 

St: 



^ 
% 



OJ — 

a/5— 

0^ — 

0.8 — 

0,4 — 

0.6 — 

0.6- 

OJ- 

0,8- 

0.9- 

1.0 — 

1.6 — 



8 



8.6- 

4 • 
4,6- 

8 

7 ■ 

8 • 

9 • 



10 



15 



20 



80 



40 






§ 



I 



5 



3 



60 



60- 

70- 

80' 
90' 



too 



J60' 



I 



Fi«. 85. Obast vob DBTBBifnnifa RasiaT^Ncs of Pipbb to Flow or 

(fim Hm$m^Wittimm» Formula, e - 100.) 



#0^ 



400 



S 

I 

I 



2 

5 

I 



S 



^ 



I 



':"! 



^ A . , ( 



—J 



PUMPS' 



385 



A^prozimftte Cost of Pumping Water. 

Triplex pow^ pump and steam engii^e, IH to 5 lb. of coal per hor8epowa>hoiur. 

Triplex power pmnp and gasoline engine, 1 pint of gasoline per horsepower-hour. 

Triplex power pump and oil engine, 1 pint of oil per horsepower-hoiur. 

Triplex power pump and gas engine, 10 to 20 cubic feet of gas per horsepower-hoiur. 

Triplex power pump and electric motor, IH to 4c per 1,000 watt-hours or Ic to 3c per 
lior8Q>ower-hour. 

Bmall steam pumps, about 25 lb. of coal per horsepower-hoiur. 

Large steam pumps, compounded, about 13 lb. of coal per horsepower-hour. 

PulscMneters, about 67 lb. of coal per horsepower-hour. 

InjeoUMTS and inspirators, about 100 lb. of coal per horsepower-hour. 

Note: The motcHr figures cover total power cost. The others cover fuel consumption only. 

Chart for Flow of Water in Pipes. The values obtained from this chart (Fig. 35) are based 

upon the UazerirWiUiama formula— 

0.11 /;i\ 0.14 0.1S 

v = er fyj XIO 

_- diameter 

where v is the velocity in feet per second, r is the hydraulic radius = ; m feet, h the 

4 

friction head, and I the length of piping; c is a constant depending upon the roughness of the 

XHpe and, upon the hydraulic radius. 

The formula can also be written 

^ 7 147.86 Q \i« 

where ^ is, as b^ore, the friction head in feet for I — 1000 ft., Q is the water quantity in gallons 
per minute, and d is the diameter of pipe in inches. 

The chart is based upon a value of c » 100, which is mostly used and considered safe for 
ordinary conditions. 



Fcnr other value of c the figure obtained from the chart should be multiplied by K 



( 



loqy- 



For information regarding coefficient c for different kinds and size of pipes, and also value 
of K for different values of c, see table below: 



Sbbop 
FtPB, In. 


2 to8 


4 


6 


6 


8 


10 


12 


16 


20 


24 


80 


86 


42 


48 


64 


60 


e 


K 


Condition of Pipe 


Year of Service for Cast-Iron Pipe 


140 


0.64 


Very imooth and 
■traight, braai, tin, 
etc 


00 



4 


00 



4 

• « 

14 

■ • 

28 


00 



4 

• • 

16 

• • 

80 
55 
95 


00 



6 

10 

16 

• • 

88 
62 

• « 


00 



6 

10 

17 

• • 

85 
68 

• • 


00 


5 
10 
17 
26 
87 

■ • 
• • 


00 


5 
11 
18 
27 
89 

• • 

• • 


00 



00 



00 


6 
12 
19 
80 
48 

• • 
m • 


00 


6 
12 
20 
80 
44 


00 



00 



00 



6 

12 

20 


00 








180 


0.616 


Ordinary straiffht, bras 
or tin 





120 


0.716 


Smooth, new iron 


5 
11 
19 
28 
41 

• ■ 

• • 


5 
11 
19 
29 
42 

■ • 
• • 


6 
12 
20 
80 
46 


6 
12 
20 


6 


110 


0.84 




• • 


12 


100 


1.0 




18 

• • 


20 








90 


1.21 




80 
46 


81 
46 


81 


80 


1.61 


Old iron 


26 


47 








60 


2.68 


Very rough 


46 

75 


50 
87 


• • 

• • 




• • 

• • 


• • 

• • 


• • 








40 


6.46 


Badly taberculated 


• • 



00 Indica t e! the very beet eaat-iron pipe laid perfectly etraight, and when new. 
In^catea good new caat-iron pipe. 



CHAPTER XIII 

STEAM CONDENSERS 

The primary object in operating engines or turbines condensing is for tiie purpose of ob- 
taining a greater amoimt of useful work from a given weight of steam supplied than otherwise 
results when the machine is operated with atmospheric exhaust. The obvious result, when a 
condenser is added, is a saving in fuel 

Referring to Fig. 1 and assuming that the expansion of the steam in the engine or turbine 
is adiabatic and is carried down to the back or exhaust pressure, the energy ccmverted into work 
is the difference between the heat content at the beginning and end of expansion. 




Fig. 1. 



cycle.) For an initial absolute pressure pi « 160 lb. per sq. in. and terminal pressure (atmos- 
pheric) pt — 14.7 lb. per sq. in., which is also assumed as the back pressure, the heat equivalent 
of the work obtained from one pound of steam is ii — ii = 1196 — 1022 s 174 B.t.u. (MoOier 
diag r am) represented by the area abqf. Now assume that by the addition of condensing i^>- 
paratus the back pressure is reduced to pi s 2 lb. absolute (corresponding to a 26" vacuum). 
The heat equivalent of the work obtained from one pound of steam for the same initial pres- 
sure is ti — it » 1196 — 907 = 289 B.t.u. as represented by the area abcde. This repre- 

(289 \ 
— — 1. j = 0.66 

or 66% represented by the area fcde. 

The theoretical steam consumption for the two conditions is: 



2546 

174 

2546 

289 



14.63 lb. per i.hp.-hr. non-condensing, atmospheric exhaust. 

8.81 lb per i.hp.-hr. condensing, 26" vacuum. 

586 



STEAM CONDENSERS 387 

Hie Uieoretical gain in economy or reduction in the steam consumption or water rate, 
for the stated conditions, by operating condensing is: 



(l.-^) = 0.398 or 39.8% 



The per cent reduction in the steam consumption of an actual turbine is about the same, 
as the water rate of the actual turbine varies in approximately the same ratio as that of the ideal 
Bankine engine. 

Example. The steam consumption of a certain 300 kw. turbine operating non-condensing on dry 
ntorated steam with an initial pressure of 165 lb. absolute is 38 lb. per kw.-hour, with atmospheric 
exhaust. Based on the above statement the steam consumption, when operating condensing with a 
26" vacuum, should be about 38 — (0.398 X 38) » 24.1 lb. per kw.-hour, which |8 approximately 
the result obtained in practice. 

The above gain in economy is, however, not a net gain in the fuel consumption of the plant, 
as api»oximately 3 to 10 per cent of the steam used by the main units must be allowed for oper- 
ating the condenser pumps, which reduces the apparent gain to an actual gain of approximately 
30 per cent fcH* the conditions of operation stated. Against this apparent gain must be charged 
&e fixed charges of the condensing equipment, cost of pumping the water required, etc. 

In <»xler to obtain a correct comparison between engines and turbines operating with and 
without condensers, the economy curves for the size of units being considered should be con- 
sulted. It is found in practice, on account of the excessive size of low-pressure cylinder required 
to accommodate the large volume of steam at very low pressures, that a 24 to 26" vacuum is 
a inactical limit for reciprocating engines. 

" The high cost, internal friction and condensation losses involved with the utilization of 
the last few inches of vacumn more than offset the amount of energy which might otherwise 
be gained." The steam turbine, on the other hand, is not limited by any such considerations and 
therein lies its greater superiority as the maximum degree of vacuum commercially possible may 
be utilized, with its accompanying economy. The decrease in the water rate of turbines due 
to an increase in vacuum is given in the Chapter on "Steam Turbines.'' 

The ciurve, Fig. 2, shows graphically the per cent increase in efficiency of the theoretical 
Ramkine engine when operating with various degrees of vacuum over that of non-condensing 
(^>eration with atmospheric exhaust. Figs. 3 and 4 show the effect of condensing operation on 
steam turbines of 300 kw. rated capacity from actual tests. 

Measurement and Degree of Vacuum. Pressures below atmosphere, in steam engineering 
practice, are ordinarily measured and stated in inches of mercury. This is the height of a colunm 
of mercury supported by the difference in pressure between the barometric pressure and the 
absolute pressure exbting within the exhaust pipe or the condenser. 

The actual absolute pressure ^, measured in inches of mercury, is then the difference be- 
tween the barometer reading h^ and the manometer or vacuum gage reading hg or ha » ^ — ^. 
The actual absolute pressure Pa measured in lb. per sq. in. b: 

Pa = 0.491 ha = 0.491 {hf, - hg) 

(1 inch mercury « 0.491 lb. per sq. in. temperature of mercmy 32® F.). 

Thus if the barometer reading is 29.8'' and the vacumn gage reads 26" the actual absolute 
pressure is 29.8 — 26 or 3.8" or p« = 0.491 X 3.8 = 1.866 lb. per sq. in. 

It is customary to refer all vacuum readings to a 30" barometer and in condenser calculations 
a 30" barometer is assumed, although in practice the actual reading in any place fluctuates 
considerably, due to the changes in atmospheric conditions. 

Free air is never absolutely free from the presence of water vapor, and as previously explained 
in the Chapt^ on "Air Conditioning, etc." in Volume I, the barometer pressure is the sum of 



388 



POWER PLANTS AND REFRIGERATION 



the partial vapor pressure and the air pressure (xnresponding to the temperature. As 
vapor is less dense than air the more vapor present in the mixture the leas will be the 
metric pressure. 

Let kg » reading of vacuum gage, in. of mercuiy, temperature of mercury, 58.4* F« 

h^ » reading of barometer, in. of mercuiy, temperature of mercury, 58.4* P* 
h^ — hg >» absolute pressure of the mixture of air and vapor, in. of mercury. 




FiQ .2. 



8 10 12 14 le 18 20 22 24 26 28 90 

Yaojum la tnctus of Heooufy 

EmciBNOT CUBTS OF THS PBBFBGT ENOmS. 



Then 30 — (^^ — h^ = the vacuum in inches of mercury referred to a 30^' baitHneier. 
The mercury column oorrecti<m for any change in temperature may be approximated by the 
foUowing equation for both the barometer and gage: 




2 4 e 8 10 12 14 16 18 20 22 24 26 28 30 

Vacuum In Inches 

Pio. 3. Bftbct of Yacuum on the Steam Ck>NSUMFTiON of a 300 Kw. Pabsonb Tuxbimb. 



h a observed height of the mercury column at temperature t. 
kg » height corrected to temperature ('. 
*; - A [1 - 0.000101 (( - (')l. 



/ 



STEAM CONDENSERS 



389 



TIm following degrees of yacuum referred to a 90" barometer are ordinarily used in oon- 
calculationB and in praotioe: 

TABLE 1 



, 


Vaemim 


fS^fipl* ft^fjpftf^tAtlfltf fiOgilMt . . , . , 


24'' to 26" 


CTfiafM^MvmftM WMMntM k^ftf J nV flDClDMI ..•...•..........>...•.••..•>•••>•.................•.... 


26" 


^fJ^^TrlinMt^^^v^S!^. 


28" 


I.'JT^ i?i man tttiin tuibliwi .... 


28.6" 










It is cueUnnary praotioe to base ocmdenser oalculations on the estimated wei^t of steam 
uaed by the engine or turbine at normal load. 

Mazinmm Degree fA Vacuum Obtainable. If we assume that dry saturated steam at temp. 
if and pressure pj is flowing into a condenser which is being supplied with watery at a lower tem- 




Fio. 4. BiracT 



8 10 12 14 le IB 20 22 ^4 28 28 30 

¥icuum lo Utthe9 

OF Vacuum on tbb Steam Consumption of a 300 Kw. Da Laval Tuiboib. 
Initial ProKure. 150 Lb. Oage Dry Saturated Steam. 



perature it, in suffieient quantity to o(mdense the steam, theoreHcaUy the final temperature of the 
ooding water (t may be equal to £«, or, in other words, the temperature t, and ocmsequently the 
preasure p« or Taouum maintained will depend entirely upon the final temperature of the con- 
densing water. If, tor example, the final temperature of the condensing water tt is 95* F., in the 
theqretioally perfect condenser the steam temperature t, will be equal to 95* and the condenser 
preasure p, » b.815 lb. per sq. in. absolute corresponding to 30 — 1.659 or 28.34 inch vacuum. 

PraoUcally, the above condition is not fulfilled in a condenser. 

The observed aheohite pressure p^ in a condenser, due to the presence of air mixed with the 
Tmpot, is the sum of the partial vapw pressure p, and the air pressure p«. That is, p^ «■ a + p« 
aoocffding to DaUonU law. Chapter on "Cooling Ponds and Towers." 

The actual temperature of the steam t, (the observed temperature of the mixture) is always 
lower than the temperature (« corresponding to p^ or vacuum maintained, the actual variation 
depending upon the amount of air present in the mixture. 

The ratio of — may be assumed at approximately 0.90. Assuming that a 2S" vacuum 



390 POWER PLANTS AND REFRIGERATION 

(referred to a 30" barometer) corres^nding to 2" pressure or Pc = 0.982 lb. absolute preamuB as 
to be maintained in the condenser p, = 0.90 pg - 0.8838 lb. absolute. 

The temperature corresponding to the vacuum is 101.17^; the temperature ocnrespondiiig 
to Pf is 97.67^, making a difference of 3.5^ 

Types of Condensers. There are two general classes of condensers, namely: (a) jet con- 
densers, and (6) surface condensers. 

In all t3rpes of jet condensers the steam and cooling wat^ mingle and the steam is oondeoaed 
by direct contact with the condensing water. 

In surface condensers the condensing water is ordinarily passed through tubes around wliieh 
the exhaust steam is directed, the transfer of heat from Uie steam to the water being made 
through the metal shell of the tubes. 

The surface condenser is particularly adapted for installations in which it is desirable to 
return the condensate direct to the boilers when the quality of the water supply is such as to 
require special treatment before its introduction into the boilers. 

JET CONDENSERS 

A classification of jet condensers includes the following t3rpes: 

Standard Type. In this type the cooling water is generally admitted at the t<^ of a pear- 
shaped vessel, the exhaust steam also entering at the top of the chamber through an ell connec- 
tion, and flowing in the same direction (parallel flow) with the condensing water. 

The condensate, condensing water and non-condensable gases are removed from the base 
of the condensing chamber by means of a vacuum or wet-air pump and delivered to the hot welL 
Fig. 5 shows a section through this type of condenser. 

If the suction lift for the injection water does not exceed 20 feet no circulating pump for 
the condenser is required. The water is raised by the vacuum maintained, this being the usual 
method of supplying the injection water for jet condensers. 

The condensing chamber is ordinarily equipped with a vacmmi-breaker, the oflSoe of which 
is to prevent water backing up mto the exhaust pipe to the eng^e in case the vacuum pump faib 
to remove the water asiast as it accumulates. The device consists of a float, located in a chamber 
attached to the cond^ns^Deftr the top but below the exhaust pipe connection, and operates a 
valve commimicating with the atmosphere. 

When the water has risen to a sufficient height in the condenser the valve is automatically 
opened and the vacuum broken and the exhaust steam escapes through the injection pipe and 
the pump to the hot well. 

This t3rpe of jet condenser is principally used in connection with reciprocating engines where 
the vacuimi carried rarely exceeds 24" to 26" of mercury "referred to a 30" barometer. The 
main advantage of this type of condenser is its relative low first cost. It is not suitable for steam 
turbine installations where a comparatively high vacuum is essential. 

The top of the condenser must not be more than 20 feet above the level of the supply of 
Injection water for condenser. 

The injection pipe should be full size of injection opening at top of condenser. If this pipe 
is over ^ feet long, use a larger pipe, and for long distances the pipe should be stUl further in- 
creased in diameter. Place a strainer on the end of injection pipe. The injection water should 
never be supplied to condenser imder pressure. 

The cylinder drains of main engine should connect into the exhaust pipe; the cocks should 
be carefully ground to insure tightness. 

Rectangular Rain Type. In this type (Fig. 7) the injection water is introduced at the top 
near one end, into an extended trough or pan, from which it overflows through numerous Aari 
tubes, falling into a second pan provided with similar overflow pipes and finally into the lower 
part of the shell and thence to the vacuum pump. 

The steam enters through the opening in the left, passes horizontally to the ri^^t through 



STEAM CONDENSERS 



Fio. S. Standard Ttps Jbt Condbnbkr. 



Pia. 8. Inbtaliatioh of Jr Condbhhis in CasivscnoK with Crom CoMFOinni EKonn. 



392 



POWER PLANTS AND REFRIGERATION 



the shower of water, ascends to the second level, passes to the left through the ui>per shower, and 
finally all that is left of the non-condensable gases (air) and contained vapcH' passes h<MiioQtall7 
to the right, and over the entering cold water, at the top to the dry vacuum pump suctioQ 
connection. 

TABLE 2 

STANDARD JET CONDENSERS 

Combined jet oondenaer and wet yaeuum pump. Type of pump — double acttng limplex. 

(referred to 30" barometer). Steam preemre at pump — 100 lb. c^ce. 



Capadi^i Steam Condenaed 
per Hour with Cooling 










Net Priee 


Diam. 


Diam. 




Total 
Weight, 
Pounda 


at Factory 


Water at 


Steam 

Cyl. 

In. 


Water 

CyL 

In. 


Stroke 
In. 






Plain 




70* 


80« 


Fitted 


400 


826 


4 


6 


6 


460 


$186 


$148 


1.000 


800 


4 


8 


7 


760 


164 


180 


1,600 


1,860 


i^ 


9 


10 


1,600 


260 


282 


2,000 


1,680 


10 


10 


1,800 


815 


842 


8,800 


2,700 


7 


12 


12 


2.060 


816 


848 


8,400 


2,800 


7 


12 


16 


2,400 


888 


868 


4,000 


8,860 


7 


12 


18 


2,760 


860 


898 


4,400 


8,660 


7 


14 


12 


8,000 


860 


898 


6,600 


4,660 


8 


14 


18 


8,660 


426 


476 


7,200 


6,000 


8 


16 


18 


4.100 


660 


680 


9,100 


7,600 


10 


18 


18 


6,800 


606 


675 


12,000 


9,900 


ii 


20 


24 


8,000 


760 


887 


14,600 


12,000 


22 


24 


8,200 


860 


950 


17,200 


14,260 


12 


24 


24 


8,400 


940 


l/)60 


20,200 


16,760 


12 


26 


24 


9,000 


1,000 


1,126 


26,900 


22,260 


16 


80 


24 


12,000 


1.826 


1,466 


29,400 


24,860 


16 


80 


80 


18,000 


1,426 


1.675 


88,400 


27,700 


16 


82 


80 


14,000 


1,626 


1.805 


42,800 


86,600 


18 


86 


80 


16,000 


1360 


2.090 



VERTICAL PUMPS 



48,000 


40,000 


20 


40 


24 


22,000 


2.824 


8.186 


68,700 


67,000 


24 


48 


24 


81,700 


8.668 


8,950 


100,000 


88,000 


80 


68 


24 


60.000 


6^65 


6.560 


107.000 


90,000 


86 


60 


24 


68.000 


6.810 


7,000 



A vacuum-breaker is located on the right of the drawing. If the wator level should riae 
abnormally in the shell, due to possible stoppage of the circulating pump, the float is raised and 
opens a valve to the atmosphere, whereupon the inflpw of water is stopped, since the circulating 
water is brought up to the condenser from a lower level by the vacuum. The steam will then 
escape through a relief valve in the exhaust line. 

This type of jet condenser is capable of producing and maintaining a hi^ vacuum, due to 
the efficient method employed to thoroughly mix the steam and condensing water and the fact 
that the air is removed by a separate pimip. A complete installation of this type of condenser 
is shown by Fig. 7a. 

Wheeler Low-level Type Jet Condenser. The standard low-level Wheeler jet condenseni 
are shown by Fig. 8. The centrifugal condensation removal pump is submerged in the lower 
part of the condenser, and is, therefore, always primed. 

This type of condenser is adapted for high vacuum work and emplo3rs a Thyseen centrifugal 
entrainment pump, described later, to remove the non-condensable gases from the tqs of the 
ocmdensing chamber. 

Westinghottse Leblanc Jet Condenser. Fig. 10 shows a cross-section through this tjrpe (rf 
condenser. 



STEAM CX)NDENSEBS 



393 





I I 

I • 






Water 
Inlet 



' //y^jm^£iimm^^^^^ 




PlQ. 7. Whbblbb J«t Oondbnbbb. 




Fig. 7a. Installation of Wheeler Jet Oondenber. 



POWER PLANTS AND REFRIGERATION 



Fio, 8. Whkblbb Low-Levbl Tim Jbt Condbkbebs. 




FlO. a. INBTAMATION OF WBBBLBII LOW-LkVEL JET OONDENBEB. 



STEAM CONDENSERS 385 

CoMdMiMT Htad. Tlie ctxd wat«r being brought to the oondenMr inlet B is distributed 
anmnd Um entire circumference through ftF^T'^ilw <^>eiiiiig C and enters the head through helioftl 
disttibutuig nouke D. These noulee give the water a rotary motion, and break it up into a 
fine sgwsy, so it mixes intimately with the steam which enters through opening B. As the cooling 
water enters by virtue of the vacuum within the condenser, the total suction heftd to inlet B 




or WMrronfflovsa LxBbuni Low-LaviL Jrr Ooin>Bin>B. 



diould not exceed about 18 ft. While the noule* /) are of generous proptMtiooB, hand holes 
an provided so sticks, leaves or other d^ris which might be brouj^t in with the water may be 
cosily removed. 

WaUr Pump. The mixture of oondenaed steam and water falls to the bottom of the OOD- 
dnsa' and is discharged by the double suction centrifugal pump F. The pump runner, as well 
•i the statioDsry guide vanes, sre of bronie. If desired, this pump may be designed to disohsrge 
against any ottenul bead such as might be imposed by coding towera, q>ray Douks er gootfal 
millsuHdy. 



396 POWER PLANTS AND REFRICERATION 

Air Pump. (F^. 11.) While the ajr- and water-pump ninneiB are mounted on the aune 
shaft, the inlet and discharge openings are entirely separate. The air-pump runner is a sin^ 
piece of cast bronze. The bronze collector cone G and btocks H areao designed that they may be 






Fio. 11. OBoas-3EcnoN OF Aib Pma or WsenNQBouBB Lssbura OoNDKHaca. 



3 i|g 



easily replaced should wear occur owing to the uee of dirty or acidulous water. Water is drawn 
into the air pump through opening / and flon-s out through the rectangular orifice J. The 
pump runner K, rotating in the direction shown, cuts off layers of water which are thrown into 
the collector cone O. Between the successive pistons of water, layers of air drawn in through 
pipe L are imprisoned. As the specific heat of air is low and its weight small compared with 
that of the water, the air on entering the pump is immediately cooled to the lowest poeaible 
temperature. The high velocity of these water pistons is transformed into pressure by means 
of the diSuser K, so that the mixture may be disdiarged against atmospheric preasure or a Bom»- 



STEAM CONDENSERS 397 

what hitter bead, aa the local conditions may demand. If water under preaaure is not available 
for putting tbe condenser in operation, the Bt«am ejects M may be used Tor this purpose. 

Tbe advantages of this pump may be easily seen. There are no ckiee clearances nor rubbing 
surfaoee requiiing attenticm. Neither are there reciprocating parts with their attendant packing 

Owing to the use <rf watw pistons, it is obvious that the air^iandling capacity of this pump 
IB much greater than the ordinary ejectw arrangement where the air ia simply carried along by 



FlO. 13. EOBBTINO EDUCTOR CoNDBNSKB. 

ErictioD. Tlte water is discharged through a comparatively lai^ opening which will allow small 
ddnis to pass without danger of clogging. Some hydraulic pumps of this general type hare 
a very narrow discharge cq>ening, extending around the entire circumference, and as a r^ult 
much trouble is experienced from foreign matter, and it is often necessary to use perfectly clean 
watar to insure satisfactory operation. 

Yacmtni-BTeaker. This is of the float type operating the valve N which opens, in case the 
water level in the condenser raises to a dangerous height, and "breaks" the vacuum. As the 
water mtera by virtue of the vacuum in the condenser, the supply is immediately cut off. 

This type <rf condenser is particulariy well adapted for high vacuum work. For vacuum 
of 26" and under the power consumption of the Lebianc air pump is higher than that of the 
reeqirocating type. 

Bdnctor Condenser. Fig. 13 shows, in section, a type of jet condenser designed for use 
in oonjunctiiMi with reciprocating engines tor vacuum not exceeding 24" to 28" mercury. 

Ia tbe Koerting oondenser tbe exhaust steam enters with the cooling water into the coo- 



398 POWER PLANTS AND REFRIGERATION 

denung dumbo', where i3ie steam is oond«ised direct by the wster. This [AyBiaU proesB 
being completed, the water jet, united with condensed steam and the son-condensable gMM, 
has to be discharged against the prESsure of the atmosphere. This mechanical wc^ ia dene bj 
the same water jet, which, for that purpose, has to ent^ the ecwdensing chambo' in a solid jet, 
and after the steam ia oondeosed enters the discharge cone or tail pipe with such a vdocity that 



r InSTAUJHO KOEHTtNa EDVCTOB OONDBNUa. 



!s the pressure of the atmosphere, being forceful enough to expel the air. To keep the 
jet straight it is surrounded by a combiiyi^ tube, in which ports are drilled at a suitable angle, 
through which the steam from the condensing chamber enters and is condensed by tiie jet. The 
holes are cut in an angle tending to give the water a high velocity. 

Fig. 14 shows the method of installing this type of condenser. 

Table 14 pvea the dimensions and rated capacities for educttv condensNS. 

An example showing the method employed in calculating the power required to oporaXe 
eductor oxuleDBerB is given under " Pow^ required to operate ooadenser auxiliaries." 



STEAM CONDENSERS 399 

Kowtiiig Holtl-Jet Ednctor Condenser. Fig. IS shows thia tn>e of coadeiuOT in Boction 
flesiKiMd (<»' high racuum work in conjunction with eteaia turbines. A vacuum of 28" mmniry 
r«fened to a 30" barometor may be obtained with this type of condenser. The principle of 
Is the same as for the eduotor oondenser previously described, but has, instead of one 



CMitral condensing jet, a number of converging jets, meeting and forming a single jet in the 
lower port of the condensing tube. This tube is cast in one piece, and consists of a series of con- 
oentrio notiks of gradually diminishii^ bore. The steam flows through the annular passages 
between the notiks,. which guide it, so that it impii^ies at suitable angle on the condensing jets. 
The multi-jet condensers are considerably shorter than the single-jet apparatus of equal 
OKpadtj, but, in epit« of this, the area of contact between the steam and the water is greater. 



400 



POWER PLANTS AND REFRIGERATION 



Relief Valve 



Relief Valve with 
U fling Device' 



■Special Head 

for Cleaning 

and Removing Con0 




Centrifugal 

Circulating 

Pump 



Fios. 17 AND 18. Installation of Bitlklby Barometric Condenbbb. 



STEAM CONDENSERS 401 

A furtlier advaiitage is gained by the form of the coadensing tube, which in vertical section is 
AH invvrted cone. In the upper part of the tube the steam is in contact with the coldest water, 
■od oondensation is keenest, so that a greater weight of steam is condensed per unit of area ot 
contact than is the case in the lower part of the tube, where Uie water b hotter. 

The method of ipBtalling this type trf condenser ia shown by Fig. 16 and Fig. 48. 



FlO. 19. BlTI.mT BaBOUBTBIC CoNDHNBEH lFraTAI.I.ATIOH 3R0WIKO AlB SaPAHUQB. 

An example giving the power required to operate this type of condenser appears later in 
the text. See Table 21 for water consumption of this condenser. 

Barometric Condensers. In this type of jet condenser the wet vacuum pump is dispensed 
with, the removal <A water being accomplished by elevating the condenser to a sufficient height 
ti)q)roximat«ly 35" O") above the hot well. The vacuum in this case being maintained by the 
ooluma of water in the tail pipe, which must be greater than the height of a column of water 
iritich would he supported by the diffsence between the atmospheric pressure and the absolute 
pnsure corresponding to the vacuum maintained in the condense. 



402 POWER PLANTS AND REFRIGERATION 

The absolute pressure correspoDding to & 28" vacuum, aaauming no atr preemt, is I lb. pa- 
eq. in. The difference in presBure (at sea level) is 14.7 — I or 13.7 lb. per sq. Id. Anumiiig a 
temperature of water in the tail pipe of 110° F. the density is 61.89 lb. per cu. ft. Under tbese 
conditions the water will stand 13.7 X 1U/S1.S9 or 31.S ft. above the level of the hot welL 

Figs. 17 and 18 shows an installation of the Btdklej/ barometric type condenser with 
double injection pipe and air separating tank. 

The object of the contracted passage of the ejector tube is to obtain a high velocity of the 



^tetiM 






Baboubthic CoNDKNun. 

water in order to insure the entrainmeut of the air as fast as it accumulates in the condenser 
head. 

On account of the vacuum maintained in the condenser the actual head, including friction 
against which the circulating wat^" pump operates, is approximately IS to 20 feet above the 
level of water in the hot well. The amount of power required to do this is the entire amount 
of power expended with this type of condenser. 

If a natural head of water is available 17 feet or more above the leviel (rf the hot well the 
inrculating water pump may be dispensed with. 



STEAM CONDENSERS 



Via. 2t. TomLaoN Baboutbic Odndkhbeb. 



1 



404 



POWER PLANTS AND REFRIGERATION 



A vacuum of 28'^ referred to a 30" barometer is guaranteed by the manufacturefs when 
sufficient cooling water at 70° F. is available to condense the amount of steam to be handled. 
A hot well temperature within 10 per cent of that theoretically obtainable is guaranteed for the 
above conditions. 

The distance from the center line of the riser to the center line o( the tail pipe is approxi- 
mately 4' 0"; hence it is feasible to have the exhaust riser come up inside the power-house wall 
and the tail pipe on the outside when desired. 

The BvUdey condenser is provided with an air-separating tank for the larger sizes as shown 
by Fig. 19. 

There are a niunber of condensers of the barometric type, designed for hi^ vacuum work. 




WstetL 












AJr 8uetfo$ 



Separator 




■Tail Pfp9 



Pio. 22. Hblandbr Barometric Condenser. 



which dispense with the ejector tube and use a dry-air pump to remove the air, the oonnecti<Mi 
for the pump being taken off at the top of the condenser. 

In the Alberger condenser, Fig. 20, special provision is made to cool the air and cond^ise 
out a portion of the vapor mixed with the air before it leaves the condenser in order to relieve 
the vacuum pump from handling an unnecessarily large volume. 

In the TomiUon, Weiss and Helander (Figs. 21, 22 and 44) barometric condensers the con- 
densing water flows over a series of trays in order to present a large surface of water to the steam, 
taking the place of a spra3ring device. 



STEAM CX)NDENSERS 



405 



In some cases, owing to the physical layout of the ground or flood conditions to be contended 
with, it is necessary to place the hot well considerably higher than the level of the water in the 
sucticHi well. 

In connection with cooling ponds, the entire system may be operated by one pump, by 
devBting the hot well at such a height that the overflow will maintain a constant head on the 



V^cupm Styge Q 




Starting Vatve 
Regulating Valva 



L Engin e Room Floor 
I l»A» la'ii — 



ll».IMll—II.JJIJ..Uil«.^f 




^ 



Cooling 
TowMr 




FlO. 23. IlTBTALLATION OF BABOMKTRIO CONDENBBR D7 CONJTTNCTION WTFE COOUNO TOWBB. 



q>ray nozzles. Fig. 24, or, in connection with cooling towers. Fig. 23, the catch basin of the tower 
may be located so that the level of the water is approximately 18 ft. above the hot well, and 
the condenser allowed to syphon its water. In this arrangement, only one pump is necessary 
to operate the ssrstem. 

SURFACE CONDENSERS 

The modem surface condenser equipment designed to produce a high vacuum for large 
units consists of a (1) condenser^ (2) water circulating pump, (3) dry air pump, (4) condensate pump. 

The condenser consists oi a cast-iron shell in which are placed a number of brass tubes, through 
which the condensing water is circulated by the circulating pump; condensation of the steam, 
passing over and around the tub^, being brought about by the removal of the latent heat of 
the steam by transfer to the colder tube surface (condensing surface). 

Modem surface condensers are either of the double-flow or the multi-flow type. 



406 POWER PLANTS AND REFRIGERATION 

la tiie (bMiblfrdow type the w&ter ptirmrn through ODe bank of tubes Etnd retuma throu^ a 
•econd bank. (Fig. 25.) 

In the multi-flow type the w&ttx is pased back and forth through several banks of tubea. 
In the hu^er Bisee baffle or rain plates are provided ao that the water of condenatttioti from 



Fia. 2i. Instaliation of Bahoubtbtc Condenbes in CoNJOHcnaK wna Coolikq Pomd. 

the uppo- rows of tubes is not permitted to fall on the lower rows, and thereby reduce the heat 
transmission of these rows by enveloping them with a blanket of water. 

A condenoer equipped with rain plates is termed a dry lube type surface eondauer. 

The effect of the baffle or rain plates is to increase considerably the average rate of heat 
transmiBBion of the bibea producing a corresponding decrease in the tube surface required for a 
given duty. 

A surface condenser equipment, as stated by C. F. Braun, should be desigtwd to aooMuplidi 
the fdlowing results: 

" Stsom should enter the condenser, be conducted freely to all parts with least possible 
rewstance, reduced to the lowest practicable temperature (and correspondintt pressure), and cm- 
verted into water. 



Si 

I 

1 
,1 

1 



408 



POWER PLANTS AND REFRIGERATION 



"Air, a non-oonductor, ^ould be rapidly cleAred from tlM beat-traiumittiiig siufaoe^ oot- 
lected at miuble places, practically freed from entroiDed water and water vapor, and oocded to 
a low temperature for removal at minimnin volume, with conaequent least expenditure of me- 
chanical energy. 

"Ctmdetuatt should olao be rapidly cleared from the beat-tr«nsmitting BuHaces, freed from 



Flo. 3ft. WBUrDlBiaNMD OTUMBBICAI. OOHmMBBB. 




air, collected at suitable points for removal, and returned to the steam generat<» at the 
practical temperature. 

"CircjJating aaler should paaa through the condenser with least friction, deposit a 
amount ot precipitated chemicals or d£bria, and absorb a maximum amount of heat." 



STEAM CONDENSERS 409 

Rg. 20 flfaom a eondetuer f<v large inataUatMrne, destgned akog the lines indicated by tho 



Tig- 27 ehowB a combinatioD of surface condeDBcr and wet rocuum pump of the recipcoeat- 
ing type such as used in medium-aiie inatallations. 

fig. 28 ehows an airangement of a WheeUr nrfoce oondauer equipmait deeiKDed for 
a work in oonnecticm with a steam turbine, a wet vacuum pump of the rotor; type 



8UBFAC> OONDI 



bong employed. The vacuum pump and centrifugal circulating pump in tbia inataUation are 
both driven by a vertical high-speed engine. 

Fig. 29 diowB a typical Wettin^fhoute surface condenser installation in which a separate 
air pump is anployed to remove the air. 

TliB air pump and condensate pump are of the LeUanc type as for the Leblone jet condenser. 
Heat Transfer in Sniface Condensers. The transfer of the heat in the exhaust steam to 
the eo<ding water in a surface condenser is dependent upon a number <A nmditiona, each of which 
<ffect the result. 

The expeoiments of Oeo. A. Omk, Trans. Am. Soc. M. E., Vol. 32, are in gedcnl ooaaidered 
to provide the most reliable data available on the heat transfer through condenser tubes. 

lite following matter, including the diagrams, has been taken bom a p^>er by C. P. Bratm, 
Trans. Ara. Soc. M. E., LQIS. 

the transfn of heat from the steam, through the condenser tube, to the cooling wat«r may 
be Rpreaented by the following formula: 

Let K, » heat transfer from steam to tube per aq. ft. per hour per dep«e difference between 
the tube surface and mean temperature of the circulating water. 
Kg — heat transfo' from tube surface to cooling water per aq. ft. per hour per degree 
difference between the tube surface and mean temperature of circulating 
water. 
C — conductivity of the tube per sq. ft. per hour per degree difference between the 
inside and outside surface temperatures of tube. 



POWER PLANTS AND REFRIGERATION 



1 



[7 — B.-t.u. Iransmitted from steam to water per sq. ft. per hour per degree difference 
between the steam ^d mean temperature of the circulating water. (Unit 
heat bansmiasioD.) 

t^ > mean temperature difference between the circulating wator and steam. 

I, — temperature of exhaust steam correaponding to the vacuum. 



FlO. 39. WCSTINOHOnaE SOKTACE CONDENBCa IHSTUXATIOM. 
1 



u = 



_ The factors K, and K^ are dependent upon the condition of the surface of the tubea u 
upon the velocities of the media from and to which heat ta being transferred. 

The mean temperature difference 1„ between the circulating water and the ateam I, 
generally assumed to be reliahly stated by Gm^iofs formula: 



tm~- 



STEAM CONDENSERS 



411 



For general purposes it is (»tliiiarily considered sufficiently accurate to use the arithmetical 

ti + tt . 



as calculated from the formula, <« »■ f« — 



in which (s « final temperature of cir- 



culating water, h «> initial temperature of water and t, » temperature of the steam cor- 
responding to the vacuun^ 

Condensing Surface Required: 

Let W » weight of steam to be condensed per hour. 

t^ B mean temperature difference between water and steam. 
U ^ coefficient of heat transmission. 
Q « total heat to be removed by cooling water per lb. of steam condensed (usually 

assumed as 1000 B.t.u. in all cases for simplicity). 
S « square feet oi cooling surface. 

WXQ 

^ t^xu : ; ^^> 

Tlie values of U, Q and t^ may be taken from the diagrams Figs. 30 and 31. 

The curve A, Fig. 30, is based on the tests by Orrok on clean tubes, the curves C and D are 

1000 

fif: 900 

"Z • 800 

•5 ^ 

i %soo 

&§400 
I i.300 



200 





































y 




■^ 










^ 


y 


ft»»€ 


^ 


^ 








A 






5^ 


^ 


^ 








J 


y^ 


^ 


^ 


i^"^ 












0\ 


/^ 












' 






/^ 






' 




























) 



/ 



& 



8 9 10 



Velocity of Water through Tubes In Ft. oer Sec. -"K" 
Pig. 30. Heat Transmission — ^Velocitt Ourvi 



ueoommended by Broun as limits f or ^ to be used in the design of modem dry-tube surface con- 
densers using copper alloy tubes ^" to ^" inside diameter as the maximum. 

Tlie velocity of water through the condenser tubes in practice Varies from 2 to 6 ft. per sec. 
the average being about 4 ft. per sec. or 240 ft. per min. 

In using the curves Fig. 31, it is assumed that the values f or r/ are to be appliied to modern 
^jrpe dry-lube condoisers designed along the lines of the condenser riiown b}^ Fig. 26. 

In the standard type two-pass surface condenser it is usual practice to limit the value of U 
to approximately 300 B.t.u. for water velocities of 4 to 5 ft. per sec. 

The C. H. Wheeler Mfg. Co. uses a coefficient U varsring from 300 to 400 B.t.u. 

Biafliple. Required the amount of cooling surface for a dry-tube tsrpe surface condenser to be 
attached to a 2000 kw. high-pressure turbine. Assume water rate 18 lb. per kw.-hour. Initial temp, 
drcolating water 70* F., 2S" vacuum referred to a 30^' barometer, temperature corresponding to vacuum 
101* F. Final temperature of water 101 — 6 ■> 95* F. Velocity of water through condenser tdbes 
5 ft. per sec C^ ■■ 640 B.tu. for safe design, eurve D, Fig. 30. 



POWER PLANTS AND REFRIGERATION 

B temneratuTB differanoa, eqiutiOD 2 or Fig. 31, >•; 
95-70 



1 



The (lesm to b« oondeniMl per hour I«: tf — 
The heat to be removed per Ih. of at 

. ,. , « 38000 X 
required, aquation a, u therefore: 5 — — -— 



The tab* Burtaea 



- - 4,386 eq. ft. 



If the ooDdeiuer ia to be of the Btandard two-pan type a value of [/ — 325 mar ^ naed. The 
lube MifaM reiiuind will be: 

Ttmperature, Org. Fahr, 



110 i 
200 100 



d. J « 
^100 ^ SO "3, 



FlO. 31. ClTRTEa FOB SOLtmON OP EQUATION <3). 

„ 38000 X 1000 _„„ 



15.2 X 325 
Compara the latter figure with the data in Table S. 



STEAM CX)NDENSERS 



413 



Tlie oondenger oooling surface for a list of recent turbine installations follows: 







TABLE 3 






Kw. Tttrfoine 


Square Fnt 
TubeSarfaee 


Square FmC Tube 

Surface Per 

Kw. 


Kw. Turbine 


Square Feet 
TuoeSurfaee 


Square Feet Tube 

Surface P«r 

Kw. 


7,500 

8.000 

8.000 

10,000 

12.000 


25,000 
28,000 
18,000 
22,000 
26,000 


8.84 
2.88 
2.26 
2.20 
2.08 


14,000 
14,000 
16,000 
20,000 


18,000 
25,000 
25,000 
82,000 


1.29 
1.79 
1.87 
1.80 



Amount of Cooling Water Required for Condensers. Condenser calculaticms are haa&d 
on the assumed observed vacuum and, as was previously shown, the actual temperature in the 
condenser, owing to the presence of air in the system, will be somewhat below the temperature 
ocMTeqxmdIng to the observed vacuum. 

It is theref<H« impossible for the final temperature of the oooling water it to rise to the 
temperature ig onresponding to the observed vacuum, and in practice, owing to the inefficiency 
of the heat transfer between the water and steam, condensers are designed for a difference In 
temperature (tg — <s) (^ 5 to 15 degs. F. This d^erence is termed "temperature head " or 
" terminal differoioe." 

In practice condensers are designed for the normal load of the machines to which th^ are to 
be connected. The initial temperature ti of the cooling water is that corresponding to average 
summer conditions. If the source of supply is a stream an initial temperature of 60^ to 70* 
is ordinarily assumed. 

If a cooling tower is to be used in conjunction with the condensing system an initial tem- 
perature of 80* F. may be assumed. 

The amount of condensing water required may be i4>proximated by the following formula: 

Let Vg » latent heat corresponding to the vacuum desired or absolute pressure p«. 
Qg » heat of the liquid corresponding to pg. 

U^Te + qe 

X ■> assumed quality of the exhaust. 

■> 0.90 to 1.0 for preliminary calculations. 
qx « heat <A the liquid corresponding to the temperature tg of the condensate. 
ii wt initial temperature of the circulating water. 
it » final temperature of the circulating water. 

(s » (t for jet condensers and q^ a qt, heat of liquid corresponding to U, 
w » pounds of condensing water per lb. of steam condensed. 

«rc + g^ — g, « w (<i — ti). 

Assuming « » 1, then to -■ for surface condensers and w -■ for jet condensers. 

ti — <i /i — h 

It is usually considered sufficiently accurate to assume the value of the numerator as equal 
to 950 B.t.u. when the steam supplied the prime mover is dry saturated and 1000 B.t.u. when 
the supply is moderately superheated. 

For more refined calculations the data given by Table 4 may be used. 

The curves. Fig. 32 (G. H. Wheeler Mfg. Co,)^ are convenient for quickly determining the 
values of w, the weight of condensing water required per lb. of steam, for various conditions of 
operation. 

Bsample. Required the weight of oondensiiig water per hour for a 1000-kw. turlnne* the water 
rate of idiioh is 15.5 lb. per kw. when operating with a 2S" vacuum, p^ * 1 lb. absolute. Initial tem- 



I 



414 



POWER PLANTS AND REFRIGERATION 



perature of oirculatixig water assumed as <i » 70^ F. Assumed difference in temperature between 
temperature corresponding to the vaouimi U and the final temperature of the condensing water tt equal 
to 15 degs. oTti » 101.76 — 15 » 86.76^rc » 1035.6, «£ »69.7. Assume x » 0.95 and tg ^ tc — 5 or 
96.7*. Ox - 64.7. 

TABLE 4 



Type of Condenser 


Vacuum 

Ins. 
Mercury 


ix 


<s 


Terminal 


Standard type jet oondenaen with wet air pump 

Modem tvoe iet condensers for high vacuum 


24" to 26" 

28" + 

27" to 28" 

28H" + 


tx't,^ 
U-IO 
tc-2to6 


U-1BU>20 
U- 2to6 

<c-2to8 


If- <x -15 to 20 
U-tx^ 2to 5 


Ordinary sunaee condensers with wet vacuum pump; 
Tnediumnriae installations 


Ic-lt- lOtolS 


Hi^i-aade surface condensers, multi-flow, with both wet 
and dry vacuum pumps for large installations 


U-tt' 2to 8 



35 40 45 50 55 60 65 70 75 80 86 90 95 100 106 W liS 120 ^ 




35 40 45 50 55 



60 65 70 75 80 86 90 96 

Temperature Condensing Water la **F. 
Fig. 32. 



m 106 UO 116 120 



To use the diagram, add the desired terminal difference to the temperature of the injection water, 
and read up from the corresponding point on the base line to the proper vacuum curve. The ratio appears 
at the left. For example, for 70 degrees water, 15 degrees terminal difference or ''temperature head'* 
and 28-inch vacuiun, read up fjrom 85 degrees, IntersectiUK the 28-inch vacuum curve at a point corres- 
pondiug to the ratio of 59 to 1. These curves are int^ded for turbines or engines using saturated steam 
at the thr6ttle. 



STEAM CONDENSERS 



415 



0.95 X 1035.6 + 69.7 - 64.7 ^^ ,^ 

• • ^ ■ :::rz — z:: ■ ^ 1d» 

86.7 - 70 

The total weight of water to be supplied condenser per min. is: 

1000 X 15.5 X 59 ,^ ^^, ,^ , „^^ . 

« 15,241 lb. or 1,830 gal. 

60 

The table given below shows roughly the d^pree of vacuum which is commercially obtainable 
with varying initial water temperatures, without imdue expenditure of capital or power, assum- 
ing that the conditions prevailing have nothing of an abnormal nature about them. (Af . W. 
KtUoQg Co,) 

Tempenttore Vacuum 

of Water Obtainable 

eOdeg. Fahr 28Hindie8 

66deg. Fahr 28 k inebm 

70deg. Fahr 28 inehes 

76 dee. Fahr 27 ^ inches 

80 deg. Fahr 27 H indiee 

86deg. Fahr 27 M indiee 

SOdeg.Fahr 27 inehes 

TABLE 5 

TESTS MADE ON WESTINGHOUSE SURFACE CONDENSERS 
EQUIPMENT CONSISTING OF CONDENSER. DRY-AIR PUMP, CONDENSATE PUMP 

AND CIRCULATING PUMP 



Rated Capacity Turbine 

Square Feet Condensing Surlaee 

Barometer 

Vacuum at top of condenser — inches, mercury 

Aboolute total pr. ins. mercury 

*Afaoohifte pr. lb. sq. in. pc 

^temperature steam conreqranding to vacuum, degrees Fahr. U 

Aetoial temperature mixture at top of eondau«r U 

*Staam pr. c or responding to U — p« 

*ASr presBuro pa^fc-'Ps 

•Ratio?! 

PC 

Vacuum at air pipe connection 

Temperature eonoensate tx* 

Initial temperature condensing water t\ 

Ffaal temperature condensing water 1% 

li—ti 

^Teraubial difference <c— ^ . . . . ; 



1,600 Kw. 


2.000 Kw. 


9,000 Kw. 
20,000 Sq. Ft. 


6,000 Sq. Ft. 


4,000 Sq. Ft. 


28.99 


29.26 


80.16 


28.66 


27.20 


28.96 


1.82 


2.06 


1.20 


0.64 


1.02 


0.688 


87.2 


101.8 


84.68 


84.0 


102.0 


88.0 


0.677 


1.008 


0.668 


0.068 


0.012 


0.081 


0.901 


0.988 


0.947 


28.7 


27.2 


29.08 


82.0 


100.0 


82.0 


69.0 


84.0 


66.6 


77.0 


100.0 


78.0 


18.0 


17.0 


11.6 


10.2 


1.8 


6.68 . 



• Data supplied by the authors. 



TABLE 6 

TESTS MADE ON WESTINGHOUSE LEBLANC JET CONDENSERS 

(Weatinuhoiut Madiim Co.) 



Load in kilowatts 

Barometer 

Vacuum at condensu* 

Vacuum corrected to 80" Bar. 

Temperature of injection deg. 
Fahr. ii 

Temperature of disdiarge deg. 
Fahr. tt 

Temperatare of steam in ex- 
haust line Is 

Terminal difference (Is — (a) • • 



National Tube 
Co., Lorain, 
Ohio, 1,600 Kw 
Westinghouse 
Low Pressure 
Turbine. 



1,700 
29.30 
26.80 
27.60 

78 

106 

108 
8 



1.760 
29.8 
26.7 
27.4 

78 

106 

110 
6 



Michigan Alkali 
Co., Wyandotte, 
Mich., 1,260 Kw. 
Allis - Chalmeri 
Turbine. 



800 
29.47 
28.60 
29.08 

62 

78 

79 
1 



1,260 
29.47 
28.80 
28.83 

62 

88 

84 
1 



Ellsworth CoUieriefl 
Co., EUsworth, 
Penna.. 760 Kw. 
General Electric 
Mixed 
Turbine. 



800 
29.00 
27.86 
28.86 

64 

77 

82 
6 



800 
29.00 
27.80 
28.80 

67 

80 

86 
6 



Utiea Steam A Mohawk VaL 
Cotton Mills, Utica, N. Y., 
1,000 Kw. Westinghouse 
Turbine. 



1,000 
29.26 
27.76 
28.60 

78 

90 

92 
2 



960 
29.26 
27.60 
28.86 

70 • 

96 

96 




926 
29.80 
27.20 
27.90 

86 

102 

108 
1 



416 



POWER PLANTS AND REFRIGERATION 







TABT.E 6 1 


[Continued) 












Narragansett Electric Empire Diet. Electric Co., 
lighting Co., Providence, Joplin, Mo., 6,000 Kw. 
R. I., 4,000 Kw. Westing- Westinghouse High 
hoose iWbine. Pressure Turbine. 


Mineral Point Polilie Swin 
Co., Bfinsral Point, Wis. 
1,200 Kw. WeatinchotMO 
Bleeder Turbine. 

• 


Load in kUowatta 


2,800 
29.90 
28.26 
28.86 

84 

98 

96 
2 


8,600 
29.90 
27.70 
27.80 

84 

101 

104 
8 


8.000 
29.40 
27.40 
28.00 

42 

100 

101 
1 


4,800 
28.91 
27.80 
28.89 

77 

92 

94 
2 


2,000 
29.80 
28.16 
28.86 

69 

80 

84 

4 


8,900 
29.24 
28.00 
28.76 

67 

88.6 

86 
2.6 


476 
29.00 
26.60 
27.60 

86 

106 

108 



600 
29.00 
26.40 
27.40 

86 

111 

111 



S60 


BaroDWtor ....tttrTt.T 


29.00 


Vaeuum at condenser 

Vaeuum eorreeted to 80" Bar. 
Tempenttore of injection deg. 

Temperatuxe of diaeharse deg. 
Fanr. It 


26.40 
S7.40 

88 

111 


Temperataxe of ateam in ex- 
haust line is 


111 


Tenninal difference <Cs ~ fi) . . 






TABLE 7 

TESTS MADE ON SURFACE CONDENSERS 
iProf. R. L. TTdffMon, "Trans. Institute Naval Axeh.,** vol. 48. 1906.) 



B 
C 
D 




4.19 
6.47 
9.68 
10.94 
17.22 
18.81 
12.62 
18.80 
86.6 
27.8 



4 

r 



64.1 
41.7 
28.2 
24.8 
81.8 
18.8 
46.9 
14.2 
26.0 
14.4 



tt-ti 



16.6 
22.6 
84.2 
89.8 
80.0 
78.0 
22.0 
71.6 
86.9 
66.8 



If 



I 



i 



4,160 
6.100 
9,230 
10,660 
16,440 
18,260 
12,640 
18,660 
83,220 
26,580 



pc 



0.68 
0.98 
1.66 
2.07 
0.67 
1.96 
0.51 
1.71 
0.96 
2.16 



^ 



I 
I. 

I 



89.2 
101.2 
119.2 
127.4 

88.9 
125.4 

80.2 
120.6 
100.4 
129.0 



61.0 
60.6 
60.1 
60.0 
46.3 
46.7 
46.0 
42.6 
41.8 
41.7 



66.6 
78.2 
84.8 
89.8 
75.8 

118.7 
68.0 

114.0 
77.2 

108.0 



lOQ 



71.9 

94.1 
117.8 
128.4 

90.0 
128.9 

76.3 
121.9 
101.3 
132.0 



I 



J 

§1 



<»-«, 



22.7 
28.0 
84.9 
88.1 
18.6 

6.7 
12.2 

6.6 
28.2 
21.0 



h 

it 



U-ts 



17.8 
7.1 
1.4 

• 1.0 
1.1 

• 8.6 
8.9 

- 1.4 

- 0.9 

• 8.0 






im 



29.8 
88.2 
60.1 
66.4 
26.7 
29.6 
21.2 
28.8 
88.6 
46.6 



^1 



K 



114 
160 
181 
198 
640 
618 
607 
644 
864 
670 



if. 

h 



Vm 



0.8 
0.8 
0.8 
0.8 
4.6 
2.1 
4.6 
2.0 
4.6 
2.0 



(A) — Old type plain condenser, K-bich tubes, 4 feet long, 6 

) — ^"Contraflo" condenser. H-inch tubes, 4 feet long, 4 
(C) — Same as "B" but with separate dry-air pump. 
[D)— Same as "B," tube length 2 feet 6 indies, 
as arranged by Prof. R, C. H, Htek, 



CIRCULATING PUMPS 

The function of the ciroulatmg pump, as the name implies, is to dther circulate the condens- 
ing water through the tubes of a surface condenser — or in the case of jet condensers, lift the water 
from the source of supply to a sufficient height so that the vacuum maintained in the condenser 
may be able to complete the lift. In general the tn&utired head for the lift by the vacuum diouki 
not exceed 20 feet. 

Both reciprocating and centrifugal pumps are employed (at this purpose. Direct-connected 
motor OF steam-turbine driven centrifugal pumps are quite universally used in electric-power 
plants. 

The efficiency curve for centrifugal pumps designed particularly (oft this class of service is 
somewhat flatter than for the usual type of pump, giving a fairly good efficiency over a ooDaider- 
able range. 



STEAM CONDENSERS 



417 



TABLE 8 

JET CONDENSER TESTS 
iWhtOer Condenmr & Engin0enng Co.) 

OLD STYLB JKT CONDENSING OUTHT* 





- 


Corrscted 


Steem 


Temperature 


, Temperature 


Vaaiam 


VajiMMiii 




Vaeuam 


Temperature 


of 


of 


Correspoodiiig 


Gm» 


Barometer 


Befemd 


CorrMponding 


Inlet Water* 


Outlet Water, 


to 




toSO-ineh 


to Vaeuiim» 






EHaeharse Water 






Barometer 


Deg. 


Deg. 


D^. 


Temperature 


25.8 


29.8e 


26.44 


121 


54 


106 


27.88 


25.2 


29.se 


25.84 


127 


54 


106 


27.70 


25.0 


29.36 


25.64 


129 


54 


107 


27.63 


25.5 


29.36 


26.14 


124 


54 


108 


27.66 


26.6 


29.86 


26.14 


124 


54 


104 


27.88 


25.4 


29.36 


26.04 


125 


54 


100 


28.07 









JBT 00NDBN8BR 








27.70 


29.29 


28.42 


98 


51 


92.0 


28.49 




27.66 


29.29 


28.87 


94 


51 


98.5 


28.42 




27.80 


29.29 


28.51 


91 


51 


92.0 


28.49 




27.60 


29.29 


28.21 


97 


51 


98.0 


28.44 




27.60 


29.29 


28.81 


96 


51 


91.0 


28.54 




27.85 


29.29 


28.65 


90 


51 


87.0 


28.91 





* This eondeneer was equipped with a dry-air pump. 



Fig. 33 shows the characteristic curves for a centrifugal pi*mp designed to operate in con- 
junciioii with a barometric condenser, the rating being based on 7500 gal. per min. when 
operatiiig with a head of 29 feet. 

The moMmum measured head or lift which various degrees of vacuum will overcome may be 
determined by the following fommla, all heads measured in feet of water column: 

Let H B total available head due to vacuum. 



P* - Pi 



0.43 



- in which p^ » barometric pressure lb. per sq. in. and p« «» absolute 



pressure corresponding to the vacuum. 
Am s measured head. 

hf « friction head of pipe, valves and fittings. 
Ay » final velocity head. 

-» — in which v =* velocity of water in injection pipe ft. per sec 

0.50 Ay a loss at entrance to pipe. 
Then ^ = A^ + A/ + 0.50 A, + A,. 



" 



Bsample. Determine the maximum measured head for which a 27" vacuum (referred to a 30 
banmieter) will lift sufficient injection water at 70® to condense 42,000 lb. steam per hour. Size of 
injection pipe 12" measured length of pipe 50 ft. with 3 elbows in the line. 

For the above conditions, assuming a 10^ "terminal difference," the final temperature of the 
condensing water will be 10® less than the temperature corresponding to a 27" vacuum (3" 
mercury absolute pressure) or 115 — 10 » 105**. From the diagram Fig. 32, we find that 27.5 lb. 
water will be required per lb. of steam condensed. The volume of water required per minute will 



therefore be: 
300 



42,000 X 27.5 



60 X 62.4 
or 7.0 ft. per see., At 



309 cu. ft., or 2317 gal. Area of 12" pipe- 0.785 sq. ft. Velocity 



- 0.762. ff - 



14.74 - 1.474 



60X0.785 64.32 0.43 

friction in the Une from Table 2, Chapter on "Pumps," is 1.3 ft. per 100 feet of run. 



» 30.8 ft. The head lost by 



418 



POWER PLANTS AND REFRIGERATION 



The head loet for each elbow from Table 3, Chapter on " Pumps," is 0.62 ft. 

1.3 



.'. hf 



-I- 3 X 0.62 « 2.5 ft. 



Then 30.8 - Am + 2.5 + 1.5 X 0.762, Bolving for hm sivcs 27.2 ft. for the maximum measure lift 
for the assumed conditions of operation. 

If the problem is to determine the size of injection pipe iot a given measured head and 
vacuum it is evidently necessary to solve the equation tentatively. The size of injection pipe 
connection is ordinarily figured for a velocity of 350 to 550 ft. per min. 

Neglecting friction in the tail pipe for a barometric condenser, the following equation may 




4000 



8000 



5000 6000 7000 

Gaflons per Minute 

FlO. 33. OHARACTERISTIC OUBYBS of CBNTBIFUOAL OlRCUUkTINQ PUICP USSD IN OONNBCnON 

wrrs 50-lN. BuLKLET Barometric Oondensbr for a 4500-Kw. Steam Turbinb. 

% 

be written: —= A- — ( ) in which h^ is the height of the column of water in the tail 

2^ \ 0.43 / 

pipe above the surface of the water in the hot welL 

Example. Assuming the bame data as in the preceding problem for a barometric condenser, da> 
termine the minimum height kx required if the diameter of the tail pipe is made 14 inches. Area pipe > 

42,000(1 +^.5) 
1.07 sq. ft. Total weight of water to be handled per min. by tail pipe ■■ 



60 



- 19.950 lb. 



STEAM CONDENSERS 



419 



19.950 5.33 r» Pb — Pc 

or — — = 5.33 cu. ft. per sec. Velocity in pipe v = — - ~ 4.08 ft. per sec. — « 0.39, 

C2.4 X 60 ^ » *- J Q7 ^ 2g 0.43 

30.S, 0.39 =* Ajc — 30.8 or hx ^ 31.2 ft.; to allow for friction and other contingencies hx may be made 
approximately 32 ft. 

Loss of Head Through Surface Condensers. The lost) of head through surface condensers 
may be approximated from the following data taken from a set of curves by W, V. Treeby 
(Power, 1910). The term "pass'' refers to the number of turns the circulating water makes 
in flowing through the condenser tubes. 

Suppose the water entered the condenser at one end, flowed straight through the tubes and 
out at the other end. This would be called a one-pass condenser. Similarly, if the water, 
instead of passing through all the tubes, flowed in one direction through half the total number 
of tubes and then returned through the other half, leaving the condenser at the same end as 
that at which it entered, this would be a two-pass condenser. 

For example, with a one-pass condenser, with water flowing at a velocity of 3 ft. per second 
through ^-inch o. d. tubes, 6 ft. long, it is foimd, upon referring to the table, that the frictional 
head in the ocmdenser would be equal to 0.9 ft. 

Now if the tubes are arranged so that the circulating water would flow through half the 
tubes to one end, then reverse and return through the other half, we would halve the water area, 
double the velocity and increase the frictional head to 3.5 ft. 

TABLE 9 

FRICTION HEAD IN SURFACE CONDENSERS 

5i" O.D. Tubes 





NuMBBR OP Passes 


Lsncth 

Tube, 
Fett 


1 


2 


8 


4 


VelodtiM 
Fett per Saeond 


VdodtiM 
Feet perSeeond 


Velodties 
Feet per Second 


Veloeities. 
Ftoet perSeeond 




8 


4 


^6 


6 


8 


4 


5 


6 


8 


4 


5 


6 


3 


4 


5 


6 


6 

8 

10 

12 

14 


0.9 
1.0 
1.1 
1.8 
1.4 


1.6 
1.7 
2.0 
2.8 
2.5 


2.4 
2.8 
8.2 
8.5 
8.8 


8.6 
4.0 
4.5 
5.0 
6.6 


1.5 
2.0 
2.3 
2.5 
2.8 


8.1 
8.6 
8.9 
4.4 
4.9 


4.8 
5.5 
6.3 

7.0 
7.6 


6.9 

8.0 

9.0 

10.0 

10.9 


2.6 
8.1 
8.4 
8.7 
4.3 


4.7 
5.4 
6.0 
6.7 
7.5 


7.4 

8.5 

9.5 

10.5 

11.5 


10.5 
12.0 
13.5 
14.8 
16.5 


3.5 
4.0 
4.5 
5.0 
5.6 


6.1 
7.0 
8.0 
9.0 
9.9 


9.6 
11.0 
12.4 
13.8 
15.3 


14.0 
16.0 
17.9 
19.8 
22.3 



Note. — ^For H" tubee, multiply above flgures by 1 .4, and for 1" tubee, multiply by 0.9. 

Example. Required the total head on an 8" centrifugal circul^ng pump supplying a two-pass 

surface condenser velocity through pipe 10 ft. per second; length of line, including allowance for ells, 100 

ft., velocity through ^" condenser tubes 4 ft. per sec., length of tubes 8 feet. The meaaured head is 

the difference between the elevation of the source of supply and the hot well level, and in this example 

10« 
will be assumed as 10 ft. Velocity head in pipe = — =» 1.55 ft. 

The loial head on the piunp is equal to 5 (pipe friction Table 2, Chapter on "Pumps'*) + 3.6 
(friction through condenser) -f 10 -f 1.55 ■■ 20.2 ft. 



AIR PUMPS 

When the removal pump must handle both the air and water it is termed a XDetroir pump. 
If the air is to be removed by a separate pump this pump is termed a dry-air pump. 

In the standard type of jet condenser the removal pump handles the condensing water, con- 
densed steam and the entrained air. 



420 POWER PLANTS AND REFRIGERATION 

For Hurfuce coDdenaere operating on the wet system the air pump handles only the con- 
tlenaate and air, &nd is therefore considerably Bmallcr in capacity for an equal weight of atcAiu 
condensed than the air pump of a jet condenser. 

Fig. 34 ahowa a section through the cyhnder of Ihu Edioanls vertical siu^e-acting "wet" 
air pump uaed in connection with surface condensers. 

It is noticeable that there is only one set d valves in thia pump, namely, the diacharge wives. 
Water enters the cyhnder of the pump by gravity from the hot well of the 






Fio. 34. Edwards Am PUMF. 

the down stroke <tf the con&«haped piston is splashed through tJie p<Hls in the cylinder wall up 
into the top of the cylinder. This process partially compresses the air already above the cylinder 
and also draws more air by inspiration from the CMidenaer shell. Hie up stnJie of the pist4M 
immediately closes the ports in the cylinder wall and compresses the water vapw and air within 
the cylinder until the discharge valves are opened and the gases and water discharged. 

This pump is of the crank and fly-wheel type with the steam cylinder above the air eylinder. 
Dimensions of the various sixes of thia type of sir pump are given by Table 17. 

In modnste-eiie plants it is a quite common arrangement to drive all the pumps used 
"with the condenser equipment with a single engine, turbine or motor. 

The dry-air pump is simply an air compressor working between the pressure limits of the 
vacuum earned in the condenser (auction preaaure) and atmospheric pressure (terminal pressure). 

Reciprocating and rotary type pumps are used for this purpose, as are also centrifu^ en- 
trainment pumps, the latter being a special form of centrifugal pump. Water is disdiarged 
radi^y by the impell^ into annular nozzles surrounding the peripfaeiy of the impeller and 
entrains the air through secondary nozilcs. 

In the "wet" system the water (condensate) serves a useful purpose in sealing the piston 
and valves of a reciprocating pump against air leakage and in absorbing a portion of the heat 
generated by compressing the air to atmospheric pressure. 

"Actual teats have proven that for ordinary wet-vacuum pumps to handle the mixture of air, 
vapor, and water, and maintain even moderately high vacuums, it is necessary to cool the coo- 
lensate 10 to 15 degs. below that due to the vacuum, which of courae requires more cireulatiog 
water and wastes mcve heat from the system. Another serious objection to Uie wet-vaeaum 
system is that compressing an emulsion ol air and water is a most effective method of mixing 



STEAM CONDENSERS 421 

tbe &ir with the oonderwate to return to boilers. Centrifugal air pumpe having no clearance 
qiaae do not lose efficiency at high vacuuina, and are rapidly coming into uae, but the recipro- 
*^f"1g type still has the advantage of requiring lees power for operation." 

Fig. 35 abows a aeetion through the MMm vacuum pump for surface oondenaera. This 
pump is ad^t«d to t^Mrate oa either the wet or dry systems. When used as a dry-air pump 



to remove air and water vapor only, a Hmall atream of water is sprayed into the suction to reduoe 
the vmpor preesure and absorb the heat of compression, also to lubricate and seal the pieton. 
"nte sealing water is utiUKed as "make up" water for the boiler feed. 

When tolerating on the wet q'stem the condeoaate flows into the pump by gravity. The 
piston creatcfl a vacuum at each stroke until, near the end of its travel, it uncovers a seriea of 



Flo. 38. CaosB-aEcnov or Leblanc Am \Nn Conhens.ite Pump. 

porta located around the middle a( the cylinder iind through which the water, air, and vapors 
are drawn. 

The valves used on the discharge ends of the pumps consist of a steel or phosphor bronse 
plate coiled at one end, the other end being left flat to serve as a Sap to the valve. 

This pump ia driven hy either a direct-acting ateam cylinder (Fig. 35) or may be of the 
chink and fly-wheel type (Fig. 27). 

Ccntrtfngil Entnlnment Pumps. The ceatrifu^-.!] typo of dry-air putup, owing to the 



422 POWER PLANTS AND REFRIGERATION 

abaeace of valves, may be driven at high speed, and is supplanting the reciprocating type to » 
conaiderable extent for high-vacuum work in connection with all forma of condensere. 

The Leblane Air Pump. By referring to Fig. 38, which abowa an air and condensate pump 
mounted on the same shaft, it will be seen that air enten \be pump throu^ the pipe C. To 
start the pump in operation, high-pressure steam is turned into the connection D. The cone 



forms the annular nozzle of a ateam ejector, so that on opening the valve in the 8t«am line a 
vacuum is created in the body of the air pump. The chamber E being piped up to a source of 
water supply, is immediately filled on account of the vacuum created by the steam ejector. 
Water then flows through the distributing noEzIc F and is projected in layers through the com- 
bining passage O into the diffuser H. Between the successive layers of water, layers of air are 
imprisoned; these layers of water (on account of the high peripheral Bpec>d of the turbine wheel 
which throws them oft) have a velocity sufficient to enable them to overcome the pressure <rf the 
atmosphere and force their way out of the pump in which a high vacuum exists. The layen of 
water act like a succession of water pistons with large volumes of air between them. 

Cold water is used in the air pump; the specific heat of sir is low and its weight small com- 
pared with that of the water, and therefore the air is immediately cooled on entering the pump 
to the lowest possible temperature. 

The water discharged from the air pump is not appreciably heatod, and may, therrfore, be 
returned to the cold well. It must be ramombered, however, that in reality a mixture of water 
and air is discharged, so that in discharging to the cold well, proper provision must be made 
for separating the air from the water. 



STEAM CONDENSERS 423 

Tfce Thjftam Air Pump. Fig. 37 shows a section through this pump. The irorkmg prin- 
ciple upon which the pump u designed consisU of two continuous water fiUns, discharged radially 
through annular nozzles euirounding the periphery of two impellers supplying the necessary 
entrainment water. These water films entrain the air through secondary noziles and the mixture 
is diacharged againat the atmospheric pressure. 

The entrainment water is supplied from a tank, usually located under the pump, and is 
recirculated through the pump as shown by Pig. 33. 

AaaoDot of Air to Be Removed from Condensers. The amount of air present in water from 




determinatioas made by G. A. Orrok, Trans. A. S. M. E., Vol. XXIV, is by volume at atmospheric 

preosure as follows: 

Fresh Croton water entering heaters and condensers 4-3% 

Leaving feed-water heater at 187" 0.93% 

Leaving hot well of condenser 0.269% 

The air Ubera^«d in the condenser from each cu. ft. of feed water or condensate is 0.0093 — 
0.00289 or 0.00661 cu. ft. at atmospheric pressure. 

At the low partial preaaures existing in a condenser this amount is greatly increased in 
vohime. 

The actual amount of air to be removed by the air pump of a surface condenser is greatly 
in excess erf the amount stated above, due to air leakage into the condensing system through the 
stuffing-boxes of the engine or turbine and the joints in the exhaust piping and condenser. 

The amount, aa stated by various authorities, varies from 0.35 to 0.55 cu. ft. of "free air" 
per cu. ft. of feed water when the condensing system is tight and in good condition. 

In the discussion following, 0.50 cu. ft. of free air per cu. ft. of feed water or condensate 
will be assumed as the amount entering the condensing system from the feed wat«r and by leakage. 

The above amount corresponds to 0.50 X 0.075/62.4 or 0.0006 lb. of air per lb. of feed water 
at atmospheric pressure and 60° to 70° F. With a jet or barometric condenser, in addition to the 
amount of air entering with the condensate and leakage, there is an additional amount of air 
hliwated by the CMidenaing wat«r that must be taken into account. 

The air liberated by the condensing water amounts to approximately 2 per cent by volume 
of the water supplied. 

With an initial temperature of condensing water of 60° F. and 4" "terminal difference," 
approximaldy 26 lb. water is required to condense 1 lb. steam f(H' a 23" vacuum corresponding 
to 1 lb. per sq. in. absolute back pressure. 



424 POWER PLANTS AND REFRIGERATION 

Far the awumecl oonditkm of operation the amount of oil lifoeroted by tite ooodeosiiig water 
wilt be Approximately 0.02 X 26/62.4 or 0.0083 cu. ft., per lb. of stevn eoadeosed, measuRd *t 
Atmoepberic pressure and 60° to 70° F. This amount corTeqxmds to 0.0083 X 0.075 - O.OOOK 
lb. of air per lb. of steam condensed. 

The weight of tur to be removed, based on a 28" vacuum per lb. (rf steam oondeased for 
surface and jet condensers fof the conditioDs assumed is: 

Surface CMwdaiBers 0.0006 lb. 

Jet CMidensera 0.0006 + 0.00062 - a00122 lb. 

The volume of saturated air to be removed by an air pump depends upon the tonperatura 
of the mixture at whi^ it is removed. 

In the case of a wet«iT pump this temperature will neceMarily be that of the condensate. 
With a dry-air pump, however, it is possible to reduce this temperature to within a few degrees 



FlO. 39. INBTAUATION' OF THTBSEN AIR PtTMP IN COHJUKCTION WITH A SUBFACE CONDENBEK. 

of the initial temperature of the cooling water if special proviflion ia made in the design of the 
condenser. This is accomplished by removing the air so that it wiU be brou^t into contact 
with the coldest water in jet condensers and the coldest tubes in surface cond«iaerB. 

The volume of a saturated mutture of air and water vapor per lb. of diy air may be detfli^ 
mined by making use of Dalian's Uw and the law of perfect gsaes. 

Let t, = observed temperature of mixture. 

T, ^ I, + 460 • absolute temperature of mixture. 
Pc — observed pressure absolute lb. per sq. in. 

■p, = pressure of saturated water vapor corresponding to t, {see Tables on the Prop- 
erties of Saturated Steam, Chapter I!). 
P, ' pressure of air corresponding to t„ lb. per sq. in. 

- P. - P.. 
V, " volume in cu. ft. of a saturated mixture of water vapor and air per lb. of air. 
144 p, V, - « r, = 53.36 T,. ' 

^ G3.35T. 

Vm = — ~CU. ft. 

' P, X 144 



STEAM CX)NDENSERS 



425 



for 



Required the volume of a saturated mixture of water vapor and air per lb. of dry air 
oondenser preasure pc ^ lib, correeponding to a 28" vacuum, temperature of condensate (« » 90^. 



O.egS lb. Then p« - 1 *- 0.698 - 0.302 lb. Va » 



53.35 (90 -f 460) 
0.302 X 144 



- 675 cu. ft. 



The curves Fig. 40 were obtained by calculation similar to the foregoing and are useful in 
determining the volume of the vapor mixture to be removed by the air pump of a oondenser. 

With both jet and surface condensers the sixe of the air pump and the power consumed 
in its operation depend largely upon the temperature at which the non-condensable gases are 



not 



1000 



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10 20 30 40 50 60 70 ^0 UO 100 110 120 

Temperatore of Air Pump Suction, Degrees Fthrenbclt. 



190 UO IM 



Pio. 40. Mixture pbb Poum> of Air at Different Temperatures and Pressures. 



withdrawn. This temperature determines the weight of air in each cubic foot of the mixture 
of steam and air passing to the pump. If with a 28-inch vacuum the air-pump suction be at 
90 degrees Fahrenheit, an air pump to remove one pound of air and the steam mixed with it must 
have a volumetric capacity of about 675 cu. ft. But if the suction temperature instead of being 
90 degrees were 70 degrees, the volumetric capacity of the pump need be only 320 cu. ft. 

Assuming that the initial temperature of the cooling water is 60® F. and that the temperature 
of the mixture at the end of the suction stroke or at the beginning of compression is 90® F., the 
volume of the mixture, V^ per lb. of air for a 28" vacuum, is approximately 670 cu. ft. from 
Fig. 40. 

The volume to be removed by the air pump based on the preceding assumptions as to the 
wm^t of air to be removed will be per lb. of steam condensed: 

Surface condensers, 0.0006 X 670 » 0.40 cu. ft. measured at 90® F. 

Jet condenserB, 0.00122 X 670 » 0.82 cu. ft. measured at 90® F. 



426 POWER PLANTS AND REFRIGERATION 

The above figures correspond to 25 cu. ft. of mixture per cu. ft. of oondenaate for siirfaoe 
condensers and 51 cu. ft. for jet condensers. 

Capacity of Diy-Alr Pumps. 

Let V ~ volume of steam condensed per minute, cu. ft. 

Vg s volume of saturated vapor to be removed per minute, cu. ft. 

— V X number of cu. ft. of air per cu. ft. of condensate entering system per min. 
E = volumetric efficiency of air pump. 
D s pump displacement required per minute, cu. ft. 

E 

According to the preceding calculations and assumptions as to the weight of air entering 
the system for a 28" vacuum, t;, » 25 1^ for a surface condenser and Ug » 51 o for a jet ccmdenaer. 
With an assumed volumetric efficiency E » 0.85, D ==SOv for surface condensers and Z> = 60 v for 
jet condensers. Oebhardt gives the following values as representing current practice: 

D = 20 p to 30 » for vacua 27". 
D = 35 V to 50 1; for vacua 28". 

Example. Required the displacement for a dry-air pump to be used in conjunction with a sur- 
face condenser attached to a 2,000 kw. turbine, 28" vacuum to be maintained with 60^ cooling water. 

18 X 2000 

Water rate of turbine, at normal load, 18 lb. per kw.-hour. v = -— — - — 9.6 cu. ft. volume of 

62.4 X 60 

condensate per min.' 

Assuming the same data as in the preceding discussion, D ■> 309 or 288 cu. ft. per min. Thia 
displacement is obtained with a 8^" X 12" double-acting pump running 74 r.p.m. or with a 24" X 12f' 
single-acting pump operating 72 r.p.m. 

Capacity of Wet-Air Pumps for Surface Condensers. In this case the pump must handle 
the condensate as well as the air. Making use of the same data as given for dry-air pumps (or 
a 28" vacuum, 

Z) = 30 1> -h » = 31 » for surface condensers. 

A practical rule given by some authorities for surface condensers is: 

D = 20 1; for 26" vacuum. 
« 30 1; for 27" vacuum. 
= 40 1; for 28" vacuum. 

Capacity of Wet-Air Pump Standard Jet Condensers. In this case the pump must handle 
the air, condensate, and the cooling water. Assume 0.00122 lb. air per lb. of steam entering 
system with the feed water and cooling water as previously stated. With a 27" vacuum and 
100** hot well temperature Va = 400. 

Then 0.00122 X 400 ~ 0.49 cu. ft. volume of saturated vap<H* per lb. of condensate. 
Let V = volume of cooling water per lb. condensate, cu. ft. 
V = volume of 1 U). condensate. 

= 0.016 cu. ft. 
Vg s volume of air mixture per lb. condensate. 
Q =s total volume to be removed per lb. of condensate. 
= F -f t; -f P,. 
For a 27" vacuum and 63'' cooling water and 15^ ''terminal difference" 26 lb. water will be 
required per lb. steam condensed. 

26 1 

0.42, V = ^-r- = 0.016, r, = 0.49. 



62.4 ' 62.4 



STEAM CONDENSERS 427 

— 0.42 + 0.016 + 0.49 = 0.92 cu. ft. With an assumed volumetric efficiency, B ^ 75 
per cent, D ^ 1.23 cu. ft. per lb. of steam, or Z) = 3 cu. ft. per cu. ft. of cooling water supplied 
(D = 3 F). Average practice gives D « 3 F for single-acting air pumps and D = 3.5 for 
double-fiu^ting pumps (G^hardl) as the displacement required for the air pimip of the ordinary 
type of jet condenser used in connection with reciprocating engines for a 26" vacuum. 
Power Required to Operate Condenser Auxiliaries. Dry-Air Pumps, 
Let p« a condenser pressure absolute lb. per sq. in. 

Pt » barometric pressure absolute lb. per sq. in. / 

D SB displacement of air pump cu. ft. per min. 
Vt " volume at end of compression, cu. ft. 
W B work of compressor ft.-Ib. per min. 
n s exponent of compression ciurve. 
» 1.4 (approx.). 
Ne^ecting clearance, we have the relation: * 



W = 7 Pt D \ 1 — ^— — I ft.-K). per min, 

n-l*^' L PcD J *^ 

-3.« „»[■---§] 



In the above equation Pe, D and pt, are known or assumed and V^ is obtained from the 
relation: 

0.71 



Pc I>" = PbVfT then Vt, = n (-)" = ^(-) 



i.hp. = ir/.33,000. 

The following diagram. Fig. 41, b based on a barometric pressure p^ « 14.7 lb. sq. in. and 
one cubic foot of air and vi^wr mixture per minute. 

To obtain the expected or probable brake horsepower of air piimp, add 30 to 50 per cent 
to the theoretical i.hp. 

Szample. Required the probable brake horsepower of a dry-air pump attached to the surface 
condenser of a 2000 kw. turbine in a preceding example, D » 288 cu. ft. per min. for 28" vacuum. 
The theoretical i.bp. — 0.0177 X 288 «- 5.1. Adding 40 per cent gives 7.2 as the probable brake 
horsepower required. 

The power consumption for the wet-vacuum pump of a surface condenser may be assumed 
the same as for a dry-vacuiun pump. The extra power required to handle the water, unless 
it is to be pumped some distance to the hot well or heater, is relatively small. 

The power required to operate the wet-vacuum or removal pump of a jet condenser may be 
estimated by the following formula: 

Let Pt * absolute pressure corresponding to the vacuum lb. per sq. in. 
p» e barometric pressure. 

h » head of water in the condenser approximately 3 to 5 ft. 
H " effective head pumped against, ft. 

Pfr — Pc . 

C « total weight of condensed steam and cooling water per hour, lb. 
B » efficiency of removal pump. 

CH 



Brake hOTsepower 



60 X ^ X 33,000' 



428 



POWER PLANTS AND REFRIGERATION 



Example. Required the brake horsepower of a centrifugal removal pump used in conjunction 
with a "low level" jet condenser attached to a 2000 kw. turbine. 28" vacuum. Initial temperature 
cooling water 70° F., terminal difference 10® F. Ratio of cooling water to steam condensed, from 
curves Fig. 32, is 45. Assumed water rate of turbine 18 lb. per kw.-hour 

C = (1 + 45) (2000 X 18) = 1.656,000 lb. 



Pb = 14.7. p« - 2. A - 4 ft. H « 26 ft. 
1.656.000 X 26 



Assumed efficiency of pump E = 0.50. Brake hora^- 



fK>wer 



60 X 0.50 X 33000 



43.5. 



The power required to operate the 'Vet*' air pump for a jet condenser may be estimated 
by taking the sum of W -t Cj as in this case the pump must handle both the air and water. 



.0300 


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Visuutfijn Inches of Msrcjry 

10 12 14 16 18 20 
Aoso/ute Pressure in Inches of Mercury 



8 
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6 
24 



4 
26 



2 
28 80 



Pio. 41. Power Required to Operate Air Pumps. 



An exact estimate of the power consumption of the auxiliaries for a prt^xieed installation 
is obviously impossible. The total power consumption is, however, not a very large percentage 
of power developed by the main units and may be approximate, with sufficient exactness, from 
the data already given for practical purposes of design. 

In practice the power consumption for the auxiliaries is approximately 2 to 5 per coit of 
the power developed by the main units. See Table 10. 

The steam required for operating the auxiliaries, however, depends upon the t3rpe of drive 
selected and will exceed the percentages stated above if steam engines or turbines are used for 
the purpose. 



STEAM CONDENSERS 



429 



Ksample* Required the power consumption of the auxiliaries (circulating pump and wet-vacuum 
pump) for a surface condenser connected to a 2000 kw. turbine; estimated water rate of turbine 18 
lb. per kw.-hour. Conditions of operation to be 28'' vacuum referred to a 30^' barometer. Initial 
temperature of circulating water, 70^ F. Assumed ** terminal difference" (tc — h) » 15°. Ratio con- 
H^fMiitig water to steam » 59. Steam temperatiu^, tc »■ 102® F. Total head for circulating pump » 
15 ft. • 

Assumed displacement for wet-vacuum pump 40 X volume of steam condensed or 40 X 

2000 X 18 

« 23,080 cu. ft. per hour or Z> *» 385 cu. ft. per min. 

62.4 

The theoretical power required for Z> >■ 1 from Fig. 4 1 is 0.018. Assuming a 40 per cent loss, the 
expected brake horsepower of air pump is 1.4 X 385 X 0.018 » 10. 

The brake horsepower of the circulating pump, with an assumed efficiency of 55 per cent, is: 



2000 X 18 X 59 X 15 
60 X 0.55 X 33,000 



30. 



The total estimated brake horsepower required is, therefore, 30 + 10 ■" 40. 

AsBuming that the pumps are driven by motors having an efficiency of 85 per cent and an assumed 

40 

loos of 10 per cent in the line and transformers, the power required at the generator terminab is—— ~ -- 

0.85 X 0.90 

«• 52 horsepower. The horsepower output of the generator is 1.34 X 2000 « 2680. 

The power required to operate the condenser auxiliaries is, therefore, 52/2680 »* 0.02 (nearly), 
or 2 per cent of the power developed by the main units. The steam used by the condenser auxiliaries, 
in this case, is 52 X 18 or 936 lb. per hour. 



TABLE 10 
POWER CONSUMPTION OF CONDENSER AUXIUARIES FROM TESTS 



1^ 

CfNtdtmvfr 


Vaeaumlns. 

Mercury 
. Referred 
torn 

80"Bmr 


Initial 

Temperature 

Cooling 

Water 

I>eK. Fahr. 


Ratio of 

Cooling 

Water, 

to Steam 

Condensed 


Weight 

Steam 
Condenaed 

1 per Hour 

1 


• 

Percent. 

of Total 

Power Used 

by Condenser 

Auxfliariee 


Refwencee 






Jet 


27.1 
28.0 
27.0 
27.8 

• • • • 

.... 
28.2 

• • • • 

28.6 


82 
71 
50 
40 

• • 

67 

• • 

30 


* 46.8 
18.0 
30.0 

■ • ■ • 

1 

1 

1 

• 90.0 

1 

^ .... 

1 

1 .... 


87,600 
28,750 

1 128,000 
70,000 

; 11,200* 

1 

36,000* 

96,200 
10,260 

32,000 


2.2 
2.6 
1.1 
1.0 
2.6 

6.4 

2.5 
8.1 

4.1 


Proc. Inst. E. E., Jan., 1906. 


LeUaaeJet.... 

BaronMtrie 

Barometric 

Snrfaee 

Sarfmee 

SorfMe 

Surfaee 

Sarface 


N. W. El., Chicago. 

South Side El., Chicago. 

CiUxens' Light, Heat and 
Power Co., Johnstown, 
Pa, 

Louiaana Purchase Exposi- 
tion. 

Edison Co., Boston. 

Nashua Tight, Heat and 
Power Co. 

Los Angdes. 



* Estimated at 18 pounds per kw.-hour. 

Unless there is the equivalent of approximately 10 to 12 per cent of the total steam available 
for heating the feed water from other steam driven auxiliaries (feed pumps, stokers, fans, etc.), 
there is no gain in economy by driving the condenser auxiliaries by motors, as the above men- 
tioned percentage of the heat in the exhaust may be returned direct to the boilers by means 
of a feed-water heater. 

Example. Assume in the preceding example that the condenser auxiliaries and the feed pumps 
are steam driven. If the condenser auxiliaries are operated by a high-speed engine having a water 
rate of 35 lb. steam per i.hp.-hour and the mechanical efficiency of the engine is 85 per cent, the 

40 X 35 

steam used per hour will be: - 1647 lb. 

0.85 



430 POWER PLANTS AND REFRIGERATION 

The weight of steam used by a direct-acting boiler-feed pump, the water rate of which is 120 lb. 
per water horsepower-hour for a boiler pressure of 150 lb. gage, may be calculated as follows: 

Assume an efficiency of pump E •» 0.80 and neglect the comparatively small weight of feed water 
required by feed pump. 

160 X (2000 X 18 + 1647) 

Then -— rrrrT ■* ^7 delivered water horsepower of feed pump. 

0.43 X 60 X 0.80 X 33,000 *~ k k 

The steam consumption of the feed pump is, therefore, 120 X 8.7 or 1044 lb. per hour. The total 
steam required for the auxiliaries is tr «» 1647 + 1044 » 2691 lb. 

w 2691 

R « —— — or — — — : — -— — — 0.07 or 7 per cent of the total steam generated is used by the 

W + w (2000 X 18) +2691 

auxiliaries. 

Assuming a loss in temperature of 10^ in the condensate returned from the surface condenser, the 
initial temperature of the feed water ti may be assumed as 90** F. 

Then the final temperature of the feed water ^ is: ts - 0.9 [{xR + (1 ~ £) (h — 32)J + 32 
(Chapter on "Feed-Water Heaters"). 

With an open type heater, atmospheric exhaust % « 1151.7. 

tt « 0.9 [(1151.7 X 0.07 + (1 - 0.07) (90 - 32)] + 32 « 153*» F. 

Power Required to Operate Eductor Condensers. The following example is quoted from 
a bulletin issued by the SchuUe and Koerting Co.: 

Standard Single-Jet Edvtdor Condenser, With injection water at a temperature of 60 deg. Fahr. 
and barometer at 30 inches, eductor condensers will maintain a vacuum of 24 in. hg. column, with a 
proportion of water to steam of 25 to 1. In most instances the quantity of water used is of importance 
only in relation to the power required for working the plant, and in this respect the installation of eductor 
condensers compares favorably with either surface or jet condensing plant. 

For comparison a compound condensing plant for a 1,000 d.hp. engine may be taken. As- 
suming a steam consumption of 20 lb. per d.hp., the plant would condense 20,000 lb. of steam per 
hoiu*, and a 12" Koerting condenser, using 1050 gallons of water per minute, would maintain a vacuum 
of 24 in. hg. The condenser is 8 feet long, and with 15 feet head of water and a 2 feet long discharge 
pipe, the total difference would be 25 feet and the actual hp. required (1050 X 25 X 8.3) -f- (33.000) «- 
6.6. An efficiency of 50 per cent, can be obtained with electrically driven centrifugal pumps with full 
allowance for motor-pump and dynamo losses. The actual d.hp. required for working such a plant 
would be (6.6 X 100) -h 50 = 13)^ d.hp., or less than 1^ per cent of the power developed by the 
main engine. 

In this calculation no allowance is made for loss by friction in pipes or for gravitation flow from 
the hot well, as similar allowance would have to be made with any condensing plant. 

MvUirJei Eductor Condensers. Taking, as example, a 1000 kw., reciprocating set, using at full 
rated output 20,000 lb. of steam per hour, a multi-jet condenser, using 72.800 gallons of wat^* per hour, 
ratio 30 to 1 , would maintain 27" mercury vacuum when dealing with this weight of steam with water 
supplied at a temperature of 60® Fahr. and barometer 30". The condenser would be 6 feet long and 
the water would have to be delivered at a pressure equal to 21 feet head at the level of the inlet flange. 
The lift for the circulating ptmip would be, therefore, 6 ft. plus 21 ft. plus allowance for friction losses 
and difference of level between the pump intake and the condenser outlet flange. 

Assuming an allowance of 6 feet would suffice for these last items, the ptmip duty would be 1200 
gallons per minute through a total lift of 33 feet, representing 9.9 water horsepower. The combined 
efficiency of the motor-driven centrifugal pump should be not less than 60 per cent, and the power 
required would be 11 kw., or 1.1 per cent of the full load output of the set 

With a turbine of the same sise requiring a vacuum of full rated output of 27^" mercury a con- 
denser using 106,024 gallons of water per hour, ratio 44 to 1, would be needed, and the power required 

106,024 
would be 1.1 X ,^ ^^^ — 1.6 per cent. 

72,800 

With an exhaust steam turbine of the same capacity, using, say, twice the weight of stetun per 
kilowatt output, the power required for working the condenser with 28" vacuum would be 3.6 per cent 
of the full load output. With circulating water at a temperature of 70® Fahr., the power required, 
other conditions being as above, would represent about 2 per cent, 5 per cent and 10 per cent of the 



STEAM (CONDENSERS 



431 



full load outfNitB, and at 75*^ Fahr. about 3 per cent, 7)^ per cent and 15 per cent. If reoooled water 
has to be used for condensing, it ia important that efficient cooling arrangements be adopted, as with 
water at temperatures above 75*^ Fahr. the quantity of circulating water and the power required for 
workins the condensing plant become disproportionately high. 




73.4^ 
73>3 2 

^^^^^^^'^ '^0 '^0 ISO 130 no 90 70 

220 200 180 leO 140 120 100 80 Tbouand 
Lotd Uf Pounds of Steam per Hour 

Fig. 42. Results of Condenser Tests. 

Suiface Condenser Test The results of a test made on a Wheeler surface condenser in- 
stalled at the south works of the Illinois Steel Co, are shown graphically by Figs. 42 and 43. 
condenser contains about 6000 1-in. tubes, corresponding to approximately 25,000 sq. ft. 




10 
8 
6 
4 
2 


Fxo. 43. 



'20 22 24 26 28 30 

Dischargo Thousand of Oaf, per MIn, 

Rbsui/ts of Tests of Oentbifuoal Oibculatino Pump. 



of surface. The circulating water makes two passes through the tubes, entering at one end and 
passing through the lower bank of tubes on both sides of the center, returning through the top 
bank to the discharge. 



432 



POWER PLANTS AND REI'TUGERATION 



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POWER PLANTS AND REFRIGEftATION 
TABLE 13 

BUFFALO BAROMETRIC CONDENSERS 



A 


• 


c 


D 


B 


F 


c 


u 


/ 


7 


JC 


No. 


111 


11 


i 




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8" 


KT'-ia" 


V 


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%>f 






























































































































w 






2,000 









ff, /. J, ff.— May be eioHded 10 per Bant. 

/. J, K,— BMBd on injeetion CKpuIty lUtBd in 

Pump cuuity, to wpply injteUon, ihould bs 

Umatwi U> be required [or any Eivan goBditiona. 



Lt ibove the qoADllty Injectioa ir 









KOERTING EDUCTOR CONDENSERS (SeoFlg. IS) 














DUUHBlOin AND Ratiho 








WAm 










































""^ 


PBi Hour 






ssa 




























GmU. 


Cu-Ft. 




20 Lb. 


aOLb. 


10 U). 


A 


* 


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M 


15 


2 




IS 


10 






Sit 


y 
























^ 


















115 


























b 






































































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?Ji! 










































































!! 






































































lis 



























I of 80 d«^«* it ahould ba BbouC GO par n 

Referring to Fig. 43, it will be noted high vtuniums are obtained with cooling water M 
73 VS" temperature. 

Ths difference in t«nip«ature between the discharged coadenaing water and the steam 
ranged from 4° to 10° during the test, and the hot well temperature nnged only 2° to 9° 



STEAM CONDENSERS 



435 



below^ steam temperature. The high coefficient of heat transmission was 500 B.t.u. per sq. 
ft. per hour per degree difference in temperature with a load of 210,000 lb. of steam per hour. 

The greatest load on the condenser was 210,000 lb. of steam per hour, but it is evident from 
the test that a load of 250,000 lb. of steam per hour would have carried with 28-in. vacuum and 
70^ water. This is equivalent to 10 lb. of steam per sq. ft. of surface, and under these 
conditions the coefficient of heat transmission would have been about 600. 

Xhe coefficients of heat transmission are calculated on the basis of 980 B.t.u. per pound 
of condensate, and by the logarithmic formula for heat transmission. 

The test on the Wheeler dry-vacuum pump showed an average mechanical efficiency of 
83.9 per cent, which is remarkably high. The test on the centrifugal circulating pump is covered 
by the chart of Fig. 43. 

TABLE 15 

SURFACE CONDENSER EQUIPMENT 

(Amimed operatinK data.) 
26" vacuum referred to 30" barometer. 

Initial temperature oT water ft s-70 deg. Fahr. « Final temoerature water d »110 dec. Fahr. 

Temperature rteam 1$ «125 deg. Ffthr. Total heaa on dreulating pump »20 feet. 

Ratio: Lb. water per lb. steam » 26. 



Lib. Steam 
Coodenaed 


Sq. Ft. 
Tube 


Outside 
Diam. 




Size of 
Air Pump 


Size of 
Centrifugal 


Size of 

Engine for 

Centrifugal 

Pump 


Shipping 

Weight, 

Total 

Lb. 


Priee 
F. 0. B. 


per Hour 


Surface 


Tubes 


Single-Aetii^ 


Pump 


Fartory 


2,650 


266 


%Z*f 


3J 


4'*- 8" X 6" 


8" 


4" X 4" 


4,660 


11,280.00 


3,600 


360 


%/tt 


81 


^"- 8" X 6" 


4" 


4" X 4" 


6,460 


1376.00 


4.600 


466 


H" 


4 


"-10" X 8" 


4" 


4" X 4" 


7,100 


1,749.00 


6,200 


620 


H". 


4 


"-10" X 8" 


4" 


4" X 4" 


8,000 


1,929.00 


7,700 


776 


^" 


4 


"-10" X 8" 


6" 


6" X 5" 


9.800 


2,296.00 


9.000 


900 


6 


"-12" X 10" 


6" 


6" X 5" 


12,500 


2,684.00 


14,000 


1.410 


hi." 


6 


"-14" X 10" 


6" 


6" X 6" 


16.800 


3,498.00 


18,000 


1.800 


6 


"-14" X 10" 


8" 


6" X 6" 


20.800 


4,081.00 


22,700 


2.276 


\k" 


7 


"-16" X 10" 


8" 


6" X 6" 


27.000 


4,797.00 


28,000 


2,800 


la" 


8 


"-18" X 12" 


10" 


8" X 6" 


34,000 


6,699.00 


86,000 


3.600 


P" 


8 


"-18" X 12" 


10" 


8"x6" 


36,000 


6,487.00 


46,000 


4,600 


U" 


8 


"-20" X 12" 


12" 


8"x6" 


40,600 


7,921.00 



TABLE 16 

SURFACE CONDENSER EQUIPMENT 

(Assumed operating data.) 
28" vacuum referred to 30" barmneter 

Initial temperature cooling water <i —70 deg. Ffthr. Final temperature water k "•86 deg. Fahr. 

Temperature of exhaust steam U —101 deg. Ffthr. Total head on dreulating pump * 20 feet. 

Ratio: Lb. water per lb. of steam —60. 



Lb. Steam 
Condensed 


Sq.Ft. 
Tube 


Outside 
Diam. 


Size of 

Air Pump 

Single-Acting 


Size of 
Centrifugal 


Size of 

Engine for 

Centrifugal 

Pump 


Shipiring 

Weittht, 

Total 

Lb. 


Priee 
F.O.B. 


per Hour 


Surface 


Tubes 


Pump 


Factory 


2300 


465 


!•»/ 


4"-10"x 8" 


4" 


4"x 4" 


7.100 


$1,749.00 


3.700 


620 


Ci" 


4"-10"x 8" 


6" 


6"x 6" 


9300 


2,129.00 


4.600 


776 


«2« 


6"-12" X 10" 


6" 


6"x 5" 


11,000 


2,449.00 


6,600 


900 


H" 


6"-12" X 10" 


6" 


6"x 6" 


12,700 


2,616.00 


8,500 


1,410 


y<' 


6"-14" X 10" 


8" 


6"x 6" 


18,000 


3,696.00 


11,000 


1300 


H" 


7"-16" X 10" 


8" 


6"x 6" 


21,600 


4370.00 


18,600 


2376 


li" 


7"-16" X 10" 


10" 


8"x 8" 


28,000 


4,916.00 


17.000 


2.800 


\i" 


8"-18" X 12" 


10" 


8"x 8" 


34,000 


6,699.00 


21,000 


8,500 


%a" 
la" 


8"-20" X 12" 


12" 


8"x 8" 


39,000 


6,766.00 


27.000 


4,500 


10"-24" X 12" 


16" 


8"x 8" 


44.000 


8,461.00 


82,000 


6,400 


10"-24"xl2" 


16" 


10" X 10" 


49,000 


9,766.00 


46,000 


7,626 


H" 


12"-80" X 14" 


18" 


10" X 10" 


69,000 


18,603.00 



Ordinoiy Type Surface Condenser Installation Data. The type of surface condenser equip- 
ment ordinarily installed in the average medium-size plant consists of a (1) two-pass condenser; 



POWER PLANTS AND REFRIGERATION 



(2) wet-vacuum pump, reoiprocatiog type; (3) centrifugal circulating pump. The 
maiatained being 26" (or reciprocating engine plants and 28" tot steam tuibinea. 



FlQ. 4fi, DIMINBION DSAWlNue O 



PRINCIPAL 



TABLE 17 

OF EDWARDS AIR PUUP 



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For higher vacuum as used in large centraJ statitm work the condenser is of the dry-tvbt 
type and a dry-air pump is employed in addition to the condensate pump. 

Tables 15 and 16 refer to instaltatious for the medium-sise plant and consist 
single wet-air pump, and a centrifugal circulating pump direct-oonnected to a 



(tf. A. Slmms.) The ri 
vacuum in the tables. 



steam condensed 
condenser surface 



oof 



- 10 for 26" 



I and 5.5 to 6 for 2S" 



STEAM CONDENSERS 




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POWER PLANTS AND REFRIGERATION 




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POWER PLANTS AND KEFRIGERATION 



WSIGBT OP WATER TO CONDENSE ONE 

POUND OF STEAM FOR MULTI-JET 

CONDENSERS 




Fta. 48. MOLTI^ET CONDENBEB. (S(W Tsblo 20.) 



TABLE 20 

DIMENSION SCHEDULE OP MULTI-JET CONDENSERS 





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CHAPTER XIV 

COOLING PONDS AND TOWERS 
HUMIDITY 

Definition. Humidity is the water vapor (steam or moisture) mixed with the air. 

The maximum weight of vapor which a given enclosure will contain is dependent only upon 
the temperature (see Steam Tables) regardless of the presence or absence of any other vapor or 
gas. That is, the weight of vapor is exactly the same whether the air b present or not. 

Dalton's Law. Each gas or vapor in a mixture, at a given tempo^ture, contributes to the 
observed pm«ure 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 pi px* 
p99 etc., » the pressure of the gases or vapors corresponding to the observed temperature, then 

p = Pi + P» + Pj, etc. 

Satorated Air. Air is said to be aalttrated when it has mixed with it the maximum possible 
amount of vapor, the amount of which varies with the temperature. The vapor itself under 
this condition is also saturated (quality x = I), If the air is not in a saturated condition, then 
the contained vapor is in a superheated state. 

The acttud humidity of the air, in meteorological work, is the number of grains (1 lb. = 7000 
grains) or pounds of water vapor contained by one cu. ft. of a mixture of air and vapor at the 
observed temperature. 

The relative humidity or degree of humidity is the percentage or ratio of the actual amount 
of moisture (grains or lb.) contained by one cu. ft. of the mixture to the amount which one cu. ft. 
of the mixture would hold at the same temperature, if saturated. The condition is stated as 
80 many per cent relative liumidity. 

It simplifies calculations somewhat, if the actual humidity is considered as the number of 
-pounds of MoturcUed vapor mixed with one pound of dry air, when saturated at a given temperature 
and pressure, and the relative humidity, as the ujeight of vapor actually mixed with one pound of air 
divided by the amount of saturated vapor mixed vnth one pound of air when saturated at the same 
temperature and pressure, and expressed as a percentage. 

Example. Let it be required to find the weight of vapor carried by one pound oi air in a saturated 
mixture of air and vapor at a temperature of 60® F. and atmospherio pressure (14.7 lb. per sq. in. absolute 
at sea level). ^ 

If Pi * Absolute ];>artial pressure of the vapor lb. per sq. in. corresponding to the temperature. 
(See Steam Tables.) 
Pt » Absolute partial pressure of the air lb. per sq. in. 
p » Total or barometric pressure » (14.7 lb. per sq. in. absolute at sea level, or 29.92 in. of 

mercury), 
p * Pi + Pj -■ 14.7 at sea level. 
From the steam tables (saturated water vapor) for a temperature of 60° F., 
Pi - .26 and density « 0.00082 lb. per cu. ft. 
p% - 14.70 — 0.26 - 14.44 partial air pressure. 
From the relation PV - MRT, (Law for perfect gases). 
Where /J for air - 53.35. T - 459.6 -f 60, P - 144 X 14.44, J/ - 1 lb., 

„ 53.35 X 519.6 ,, ^, , , , ^ . 

" \aa \/ ^A aa ■ ^^'^ ^^' '*• ^oJ*"»o t>f the au*. 

•»» X 14.4:4 

441 



442 POWER PLANTS AND REFRIGERATION 



§ 



Thii also is the volume of the saturated vapor, as the air and vapor oooiq)y the fame amount d 
space. The weight of the saturated vapor is therefore: ' 

13.33 X 0.00082 or 0.01003 lb. per lb. of the air in the mixture. 

The weight of vapor per cu. ft. of the mixture is 0.01003/13.33 - 0.00082 lb. or 0.00082 X 7000 «* 
5.74 grains. The density of the mixture (1 lb. of air and its vapor) is 1.0109/13.33 or 0.0758 lb. and 
its specific volume is 1/0.0758 or 13.18 cu. ft. 

Formula for Saturated Air. (100 per cent Relative Humidity.) The operation in the 
previous problem may be expressed by a formula as follows: 

t = Temperature of the mixture degrees F. 
T = Absolute temperature = (< -f 459.6). 
P = Barometric pressure lb. per sq. ft. 

Pg s Absolute vapor pressure lb. per sq. ft. corresponding to temperature t. 
Pa *= Absolute air pressure lb. per sq. ft. 
P = P, + P« and P« - P - P,. 

V = Specific volume of air (cu. ft. per lb.) at temperature /. 
Vg = Specific volume of saturated vapor at temperature U 
Ds = Density of saturated vapor at temperature t. 

W = Weight of saturated vapor per pound of dry air in the mixture = V D,. 
Ttei PaV = ler = 53.35 (t + 459.6). 

,^ 63.35 (t + 459.6) 53.35 + tf -f 459.6) 



and TT = 



Pa P-Ps 

53.35 (t + 459.6) D, 
P-Ps 



^ 0.37 {t -f 459.6) />, , , 
or rr = — — when the pressures are stated m pounds per sq. men. 

If the weight is stated in grains and the pressures in inches of mercury, 
1 lb. = 7000 grains, 1 in. merciuy = 70.721 lb. per sq. ft. 

in which G — grains moisture per lb. of dry air, P^ = barometric pressure inches of mercury 
and Pn = absolute pressure of saturated water vapor corresponding to the temperature, in 
inches of mercury. See Table 1 for " Properties of Saturated Air/' also " Heat Exchange 
Diagram" (Fig. 4). 

Dew Point Temperature. The temperature corresponding to saturation (100 per cent 
relative humidity) for a given weight of vapor is known as the dew point. 

Any lowering of the temperature produces a contraction of volume and a partial conden- 
sation, the amount of vapor condensed being the difference between the original amount and 
the amount carried at saturation for the new or lower temperature. Air with any amount of 
vapor has a ''dew point/' as the temperature can always be lowered so that condensation must 
take place. 

The maximiun amount of saturated vapor which may be mixed with air in forming a 8aliir 
rated mixture may be calculated by making use of Dallon's law of partial vapor pressures. 

The total pressure (barometric pressure) of a mixture of air and vapor is made up of the 
sum of the partial vapor pressure (vapor tension) and the partial air pressure. 

Adiabatic Saturation of Dry Air. If absolutely dry air is passed through an insulated 
chamber containing a sponge, saturated with water, Fig. 1, it is observed that the temperature^ 
of the water will be lowered until a stationary temperature I' is reached, which is lower than 
the temperature t of the incoming air. Furthern.oro, the tcn;perature of the leaving saturated 
air will be the same as the temperature of the water. 



COOLING PONDS AND TOWERS 443 

It is pvideot that an exchange of heat must take place between the air and water, aa heat 
ia neithGr supplied by or extrnct«d from an external source. A heat tranflfer of this sort is saiil 
to be adiabatic. 

The evapmvtion ot the water takee place at the recorded temperature iA the Liquid 1'. 
Let W = weight' of water evaporated per lb. of dry air passed through the apparatus, 
determined by actual measurement. 
K = latent heat of saturated vapor corresponding to temperature of liquid or ('. 
Then 0.2411 (( — t') — B.t.u. given up by one pound of air, 

t" W = heat required to evaporate the weight of mcHstuie added to the air. 
T* W = 0.2411 <( — f) which is the equation for the adiabatic saturation of dry air. 
If the experiment wne performed with dry air having an initial temperature 1 = 75" the 
oLeerved l«mperature of the water would be (' = 46° and the weight qf water evaporated per 
lb. (tf dry air, by mesBurement, W =• 0.00056 lb. The latent beat for 46° is r* •^ 1005.6; then 



Satvkation or Dbt Aib. 

the heat required for evaporation Is 1065.6 X 0.00656 or 6.99 B.t.u. which is seen to be exactly 
the same as the heat given up by the pound of dry air or 0.2411 X (75 — 46) or 6.09 B.t.u. 

Adiabatic Saturation of a Hizture of Air and Vapor. Assume (Fig. 3) a saturated mixture 
of 1 lb. of dry air plus Wi lb. of vapor corresponding to temperature (g as obtained above corre- 
sponding to condition (2). If the temperature of this mixture is now raised to I corresponding 
to condition (3) the vapor is superheated. 

The mixture will become adiabatically saturated at temperature I' corresponding to con- 
dition (4). 

The heat given up by the superheated mbrture of I lb. of air plus Wi lb. of vapor in having 
its temperature lowered from ( to I' is 

C^ ft - O + C„ W^ {t - O B.t.u. 

Cf, •= Sp. heat of vapor at constant pressure. 

If W is the weight of vapor in a saturated mixture at temperature (' then the weight of vapor 
added to saturate the mixture adiabatically is (H" — Wi). And the heat required tor evapora- 
tion is r" (W — If])- As this is an adiabatic change, no heat supplied from an external source, 
the following equaUty exists: 

r- (W - Wi) = C„ (I - O + C^ Wi (t - f). 
The constant weight of vapor lines are plotted by adding the heat required to raise the tempera- 
ture of the mixture from saturation fi to the required temperature (. Thus for condition (3j 
add (.'^ (I - (,) + C„Wi {I -10 to the B.t.u. in 1 lb. of saturated air above 0° at temperature 
ii, coDilition (2). 



444 POWER PLANl^ AND REFRIGERATION 

The Wet and Diy Bulb Psychrometer Principles Inyolved. The actual amount of moisture 
mixed with the air under various conditions of temperature and degrees of saturation is most 
conveniently ascertained by observing the temperature at which evaporation takes place, and the 
actual temperature of the air. 

The temperature at which evaporation takes place is recorded by a thermometer, around the 
bulb of which is placed a moist cloth. This thermometer is termed the iDel bulb thermometer. 

If the spray water, through which air not initially saturated is passed, as in a humidifier, 
be simply recirculated and not supplied with heat from an external source in order to m^nf^in its 
temperature constant, and having an initial temperature higher than that of the entering air, 
the temperature of the water will soon be lowered to that of the entering air. The water wilJ 
then not be able to heat the air further, but will have its temperature lowered by any evaporation 
that may take place. The temperature of the water being lowered by evaporation, the oooted 
water will lower the temperature of the air, which, in turn, will give up some heat to the water 
by the reduction of its temperature. This heat exchange, between the air and water, will con- 
tinue until a stationary water temperature (O is reached, at which point the heat given up by 
the air to the water will just balance the heat required for evaporation. As no heat is supplied 
from an external source, it will be observed that this is an adiabatic change. 

The air leaving is then in an adiabatically saturated condition, the temperature of which is 
that as recorded by the wet bulb thermometer, as the action described is similar to that which 
takes place when air is passed over the wet cloth of the wet bulb thermometer. 

This furnishes a means for ascertaining the actual amount of moisture mixed with the air 
as given by the following method, devised by IF. //. Carrier: 

Psychrometric Meifuxi for the Determination of the Actual Weight of Moislwre per Pound ef 
Dry Air, 

t = temperature of the air degrees F. (dry bulb). 

t' = temperature of the air wet bulb. (This is the temperature at which the air 
becomes adiabatically saturate, and not the deiv-point temp.) 
i — f — wet bulb depression. 

W « weight of moisture actually mixed with one lb. of dry air at temperature L 
W ^ W — weight of moisture per lb. dry air added in order to saturate the air. 
r' = latent heat of vaporization at temperature t\ 
(W — W)r^ ^ heat necessary (B.t.u.) to evaporate (IF' — W) Vb. water at temp^ature <'. 
Cpg — Sp. heat of vapor at constant pressure (average value 0.44)-. 
Cpa » Sp. heat of air at constant pressure (average value 0.24). 
As this is an adiabatic change (no heat abstracted or added from an external source), the 
heat required for evaporation being supplied by the air and its contained vapor in lowering the 
temperature from t to t', then 

(IF' - IT) r' = Cps IF (i - O + Cp« « - O 

r' TF' - 0.24 {t - (') 



W 



r' + 0.44 (t - t') 



The relative humidity is the ratio of W/Wxt H^x being the weight of moisture per lb. of 
when saturated with vapor at temperature t. The dew point tempo^ture is the temperature 
corresponding to saturated air containing IF lb. of vapor per lb. of air in the mixture and should 
not be confounded with the loet bulb temperature. 

The determination of the actual weight of vapor in one pound of dry air is most con- 
veniently made by the use of the wet and dry bulb sling psychrofneter. 

This instrument (Fig. 2) consists of a wet bulb thermometer mounted adjacent to a dry 
bulb thermometer and so arranged that the entire mounting, which is about 15 inches in length, 
may he swung about a handle. In order to secure acciu*ate and consistent results the instrument 
should be revolved from 150 to 225 times per min. For very accurate work Carrier states 



I 
I 



eos<« 



V 



n 



<:0()UNG TOXDS AND TOWERS 



445 



that a negative correction for radiation of approximately 1.6 {)er cent of the wet bulb depression 
should be made to obtain the true depression. 

A mcNre refined type of apparatus, known as the Assntann Aspirating Psychrometcr^ makes 
use of a small fan to draw air over the thermometer bulbs at a constant rate, and in addition 
each bulb is car^ully shielded to protect it from radiation. 

Heat Kidiange Diagram and Psychrometric Chart The heat exchange diagram (Fig. 4) 
is plotted using temperatures as abecisae and B.t.u. as ordinates. The heat required to raise 
the temperature of 1 lb. of dry air from 0** to any temperature i is equal to Cpa t (C p^ *=* 
specific heat cl air at constant pressure ^ 0.2411). The dry air line having been drawn as 
shown, the satttroHon curve is plotted by adding the heat required, rWf to evaporate the 
weight of vapor mixed with saturated air, as may be calculated, to that of one pound of dry 
air above 0**, for the same temperature. The heat required to raise the temperature of one 
pound of dry air from zero to the required temperature and evaporate the weight of moisture 
added to saturate the air is known as the heat content of saturated air and is expressed by 
the formula Cp^t-hrW. 

To find the per cent relative humidity when the wet bulb reading is 66^ F. and dry bulb read- 
ing is 84^ F. The intersection of the horizontal line through 66^ F. on the saturation curve and 
the v^lical through 84^ F. dry bulb temperature gives approximately 37 per cent for the relative 
humidity. 

The dew point temperature for the above condition is found by following the diagonal 
constant weight vapor line to its intersection with the saturation curve giving 65** F-f. 

The actual weight of vapor mixed with one pound dry air is therefore 0.37 X 0.0252 (weight 
of vapor per lb. of dry air when saturated at 84° F.) or 0.000. This may be read direct on the 
saturatioi^ curve for 55° F. The dew point temperature should not be confounded with the 
temperature of adiabivtic saturation which is always recorded by the wet bulb thermometer and 
in this case is 66** F. 




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450 



POWER PLANTS AND REFRIGERATION 



CONDITIONS REQUIRING COOLING PONDS OR TOWERS 

In localities where the water supply is limited or is only obtainable at a comparatively high 
cost, the cooling water required for steam and ammonia condensers, gas and oil engines may be 
continuously recirculated when it is feasible to construct a cooling ix>nd or install a cooling tower. 
The cooling effect is obtained by the evaporation in the air of a small portion <^ the water, 3 to 
7 per cent of the amount circulated, which represents the total amount of fresh make-up water 
to be supplied. 

On account of the comparatively large evaporating surface necessary cooling ponds wiihovi 
sprays are not often used. 

On account of the excessive amoimt of water to be pumped and initial cost (A oonstructicmy 
neither cooling ponds nor cooling towers are ordinarily installed for steam condenser work 
requiring a vacuum in excess of 26 in. hg., the ordinary demand being between 23 and 25 in. 
refdred to a 30-in. barometer. 

TABLE 2 

AVERAGE ATMOSPHERIC CONDITIONS FOR VARIOUS CITIES DURING THE SUMMER MONTHS 



City 


Mban Rblativb Humidity 
Pbr Cent 


1- 

Mban Tkmpbbatukb 
Dby Bulb 




June 


July 


Aug. 


June 


July 


Aug. 


Boston 


71.6 
72.6 
67.9 
72.6 
78.7 
78.8 
77.2 
79.6 
69.7 
70.6 
72.9 
67.9 
68.2 
70.0 
46.8 
69.0 
80.1 
74.7 


71.4 
73.6 
69.8 
74.4 
79.8 
79.7 
77.7 
77.4 
67.8 
68.2 
69.6 
66.0 
66.1 
68.4 
49.0 
64.3 
84.4 
76.6 


76.4 
76.4 
71.9 
76.8 
81.4 
81.4 
78.9 
78.4 
69.0 
70.6 
71.4 
69.6 
67.6 
69.6 
44.0 
67.8 
86.8 
76.8 


66.8 

68.6 

71.2 

72.7 

78.4 

79.0 

79.6 

80.9 . 

71.1 

67.9 

66.8 

67.4 

76.1 

78.0 

66.4 

61.8 

67.0 

64.6 


71.3 
78.6 
76.8 
76.8 
81.8 
80.9 
81.8 
88.0 
74.6 
72.6 
72.2 
72.1 
79.1 
77.6 
71.8 
66.8 
67.8 
67.4 


68.9 


New York 


72.2 


PhiladdphUi 


78.8 


Wuhington 


74.6 




80.8 


Jaekaonville, Fla 


80.1 


New Orleans 


81.0 


Galveston 


82.6 


Pittsburg 


72.6 


Cleveland 


70.4 


Ch{fiip> 


71.2 


St. Paul 


69.6 


SL Louis 


77.2 


Kansas City 


76.8 


Denver . . . '. 


70.4 


Portland, Ore 


66.9 


San Franeiaoo 

liOff Anseles . 


68.0 
68.6 







COOLING PONDS 

Cooling Ponds without Spray Nozzles^ Box, in his treatise on ''Heat/' gives the following 
formula for the rate of evaporation from a pond or reservoir in still air: 

O = (240 + 3.70 (P, - P) in which 

G » grains moisture evaporated per sq. ft. per hour (7000 grains = 1 lb.)* 
t B average temperature of the water, deg. F. 
Pa -> pressure of saturated vapor in inches of mercury corresponding to the temperature L 
P = the actual vapor pressure of the air in inches of mercury. 

Actual tests of cooling ponds under average summer oonditions, wUhoxil sprays, have shown 
that ^proximately 4 B.t.u. are dissipated per sq. ft. per hour per degree diffa!ence in temperature 
between the air and the average temperature of the water in the pond. During the winter months 
the heat loss is reduced to approximately 2 B.t.u. 

The suction well should be provided with removable screens with submerged openings in 
order to prevent the hot surface water from short circuiting. The inlet pipe shouM be submerged 
at least 6 ft. below the surface, as otherwise air is liable to be drawn into the circulating water, 
which is detrimental to the vacuum in the jet tjrpe condenser. 



COOLING PONDS AND TOWERS 



451 



TABLE 3 

AVERAGE TEMPERATURE AND HUMIDITY FOR VARIOUS CITIES IN THE UNITED STATES 





Mban Valuhb fob 


Januaby I 


Mean Vauibb fob July 




City 


Humidity 
Par Cent 


Tamp. 
Dcc.F. 


Lb. Water 

per 1000 

Cu.Ft. 

Air 


Humidity 
Percent 


Temp. 
DCC.F. 


Lb. Water 

per 1000 

Cu. Ft. 

Air 


Temp. 
Cooled 
Water* 
Dec. F. 


Albmy, N. Y , . 


80.4 
76.4 
71.6 
78.9 
72.1 
82.1 

• • • • 

62.6 
47.8 
88.9 
67.4 
78.4 
78.9 
76.2 
76.8 
86.4 
74.8 
80.0 
79.7 
78.8 


28.0 
42.6 
84.0 
6.6 
27.0 
28.6 

• • • • 

29.0 
44.0 
68.6 
64.0 
20.0 
64.0 
80.6 
22.6 
88.6 
82.0 
11 .0 
60.0 
82.6 


• • • • • 

0.344 

6!866 
0.690 
0.800 

• • • • • 

0.660 

• • • • • 

• • • • • 

0.880 

• • • • • 

• a • • • 

0.470 

• • • • • 


71.9 
76.6 
69.6 
66.8 
71.4 
69.6 
68.2 
40.0 
46.0 
77.4 
76.6 
70.6 
77.7 
78.6 
76.4 
64.8 
66.1 
66.0 
84.4 
66.6 


78.0 
78.6 
77.6 
70.0 
72.0 
72.6 
72.6 
72.0 
82.0 
84.6 
71.0 
70.0 
82.6 
74.6 
68.6 
67.0 
79.6 
72.6 
69.0 
79.0 


1.01 
1.26 
1.18 
0.80 
0.96 
0.96 
0.94 
0.66 
0.86 
1.60 
0.96 
0.88 
1.67 
1.09 
0.88 
0.07 
1.16 
0.90 
0.72 
1.14 


76 


AUanfft, G« 


81 


Baltiinon» Md 


79 


Binarek,' N. D 


70 


Boston, Maa 


77 


CUeafo, in 

Clm&ttid,0 


76 
78 


Denver, Colo 


74 


Va Ptmt, TfK 


61 


Galveston, Tax 


80 


LoaAnniM, Cal 


66 


Mflwaukaa, Wia 


67 


Now Oriaana, La 


78 


Now York, l4. Y 


71 




67 


Portiand, Ora 


62 • 


9t. I^flvk, Mo, 


74 


St-Panl, Minn 


67 


San Franeiaeo, Cal 


68 


Wiehita, Kan . ... 


73 







* Probable temperature to whieh water may be eooled by a wdl-proportioned eooling aystem. 

There is nothing gained, from the standpoint of cooling effect, in constructing a cooling pond 
more than 3 ft. in depth. 

Bzample. Required the area of a co<ding pond without sprajrs for a 500 i.hp. condensing plant 
having a daily load factor of 40 per cent, the steam consumption averaging 15 lb. per i.hp.-hour. 

Assuming that the average summer temperature for the locality is 75° and a relative humidity 
of 60 per cent, initial temperature of the water from the hot well 110°, the water is to be cooled 
down to 80° and returned to a barometric or jet condenser. 

Approximately 32 lb. of injection water will be required per lb. of steam condensed to main- 
tain a 26-in. vacuum with a 15-deg. ** terminal difforenoe." 

15 X 32 X 500 X 0.40 « 96,000 lb. condensing water per hour. 

Weight of steam condensed per hour is: 

15 X 500 X 0.40 = 3000 lb. 



Then 96,000 + 3000 
110° F. 



SdtUian. Average temperature of water t = 



99,000 lb. entering cooling pond per hour at a temperature of 

110 + 80 



95° (approx.) ; P, corresponding 



to 95° = 1.650 in. hg.; P corresponding to 75° and a relative humidity of 60 per cent = 0.8744 
X 0.60 ^ 0.525 in. hg. O » (240 + 3.7 X 95) (1.659 - 0.525) » 671 grains evaporated per 
sq. ft. per hour or i^proximately, 0.10 lb. 

To evaporate 1 lb. of water from and at 95° requires the addition of 1039 B.t.u. (latent heat) 
Therefcwe Uie evaporation of 0.10 lb. will remove from the water 1039 X 0.10 » 104 B.t.u. 



per sq. ft. per hour. This corresponds to 



104 



95-76 



or 5.2 B.t.u.* dissipated per sq. ft. per hour 



*Tbii ii greater than ie generally found in practice. 



452 



POWER PLANTS AND REFRIGERATION 



per degree diflference between the average temperature of the water and the surrounding air. 
The heat to be abstracted from the water is : 

99,000 X (110 - 80) = 2,970,000 B.t.u. per hour. 

2,970,000 
Area of pond for 40 per cent load factor = — 7;;^;^ — = 28,550 sq. ft. 



Area of pond for 100 per cent load factor = 



104 
2,970,000 



= 71,300 sq. ft. 



104 X 0.40 

Basing the area on 4 B.t.u. dissipated per sq. ft. per degree difference in temperature per 

5.2 
hour increases the above figures to 28,550 X --r — 37,100 sq. ft. for 40 per cent load factor and 

4 



37,100 
0.40 



92,800 sq. ft. for 100 per cent load factor. 



The above figures correspond to: 

37,100 



500 
92,800 
500 



= 74 sq. ft. per i.hp. for 40 per cent load factor. 
= 186 sq. ft. per i.hp. for 100 per cent load factor. 



^ , ,. . • .^ 28,550X0.10 ^^ 
The per cent loss by evaporation is 100 X ^^ — or 2.9. 

Cooling Pond Test The results of tests obtained from a cooling pond located at Wampum, 
Pa., as reported by the " Practical Engineer," July 15, 1912, follow: 

It appears* that under conditions in the northern part of the United States with engines 
using 15 lb. of steam per hour per horsepower with a vacuum of 26 in., a reservoir having a sur- 
face of 120 sq. ft. per hp. would be ample for cooling the condensing water. 



TABLE 4 

HEAT RADIATION TESTS ON CONDENSERrWATER RESERVOIR 

of reiervoir, 288,000 aq. ft.; avenge depth of reservoir* 5.36 ft.; capacity of reservoir, 1,543,680 cu. ft. — 96,480,000 

pounds 



Date of Tests 



Amount of water pumped from river, lb 

Average temperature of river water, deg. F 

Average temperature of intake to power-house, deg. F 

Average temperature of tail water from condenser, deg. F 

Average temperature of reservoir, deg. F. 

Average tempermture of air, deg. F 

Average difference of temperature between water and air.i 

deg. F 

Change in temperature of r e s er voir during test, deg. F 

Steam condensed by engines, lb 

Steam condensed by compressors, lb 

Latent heat of steam condensed, lb 

Heat delivered to reservoir by engines, B.t.u 

Heat delivered to reservoir by compreosois, B.t.u. 

Heat to raise river water to average temperature of reservoir, 

B.t.u 

Heat given up or retained in r e s e r v oir during test, B.t.u 

Heat reduction in r e s e r v oir due to rain, B.t.u 

Heat absorbed bv air and evaporation during seven days, B.t.u . 
Heat absorbed oy air and evaporation per sq. ft. of surface 

seven davs, B.t.u 

Heat absorbed by air and evaporation per sq. ft. per hr., B.t.u . . 
Heat absorbed by air and evaporation per sq. ft. per hr., per 1 

deg. difference B.t.u 



Week Ending 
May 7, 1911 



10,458,966 

57.5 

72.75 

101 .36 

87.05 

61.00 

36.05 

0.25 

5.752,289 

877,204 

1024.7 

5,894,370,000 

898,871,000 

309,062,000 

24,120,000 

21,630,000 

6,438,429,000 

22.356 
133.1 

3.69 



Week Ending 
July 12, 1911 



29,050,875 

77 

91.43 

129.43 

110.00 

78.43 

31.57 

7.00 

6,433,045 

936,314 

1007.1 

6,478,720.000 

942,961,000 

958,679,000 

676,360,000 

56,700,000 

5.730,942,000 

19,899 
118.4 

3.71 



Week Ending 
Nov. 27, 1911 



4,648,140 
86 

61.71 
90.71 
76.71 
83.30 

48.41 

2.00 

6,146,148 

876,273 

1026.0 

6304,922,000 

899,056,000 

189,226,000 

192.960,000 

46,200,000 

6,564,509.000 

23,495 
139.8 

3.22 



COOLING PONDS AND TOWERS 453 

Cotriiog Ponds with Spnj NoiileB. By spraying water iuto the air, a cooling may be effected 
tfaroi^ the evaporation of a part of the nater, as in the cuee in the cooling tower. 

The total exposed surface uf the sprayed jet meets less air per pound than in the cooling 
tower, and on thig account it ie often advisable t^i epray 30 to 60 per cent of the water a second 
time before sending it through the condenger. 

Generally, ei^ay notiles of the sise knovm as 2-inch are the most economical. The 2-inch 
use acrewB on to a 2-inch outlet, the opening in the jiozzle tip being about 0.8 inch. As many 



noizlea should be provided as are needed to discbarge the entire neight of condenaing wat«r 
under a presure of not over 15 pounds gage at the nozzle. 

The noislee should be set from 8 to 10 feet apart, if 2 inch ; a greater distance if over 2-inch. 
Where a considerable number of noziles are used, it is customary to have the wat«r which is 
sprayed into the air fall back into an artificial pond one or two feet deep. 

When a number of nozzles arc in iiHe Ihe aspirator action exerted by the jets causes a current 



454 



POWER PLANTS AND REFRIGERATION 



of air to flow along the surface of the pond from the edge toward the center. This current of 
air assists, to some extent, in the cooling. 

In some few instances spray nozzles have been put along the edges of a narrow brook and 
the falling spray caught on board fences inclined 30 degrees with the ground and draining into 
the brook. 

There are several small plants where the cooling nozzles discharge on to the roof of the 
building. The extra head of water on the circulating pump, however, makes this inadvisable. 

Experiments on SchueUe-Koerting nozzles of sizes known as 3-inch, 2-inch, and 1-inch have 
been carried on at the Maasachusetta IrutituU of Tecknoloffy since 1906. 

The nozzle under test is placed at the center of a flat roof about 44 feet by 40 feet, 8k>ping 
1 foot in 10 feet, and the water caught on the roof drained into weighing tanks and weighed. 

The discharge through the nozzle is figured from the pressure shown by a gage attached to 
u piezometer just beneath the nozzle, the coefficient for each nozzle having been determined to 
three figures by exhaustive tests made in the laboratory. From the tests on the SchueUe-Koerting 
nozzles, it appears that: 

(1) The te:nperature of the water after spraying is more dependent upon the temperature 
and humidity of the atmosphere and upon the fineness <^ the spray than upon the initial tem- 
perature of the water. Therefore it is advisable to spray the water as hot as may be without 
excessive steaming. 

(2) At high hmnidity, 80 or 90 per cent, the temperature of the water may be lowered to 
within 12 or 13 degrees F. of the temperature of the air, with a total drop in temperature of 
35 to 40 degrees F. 

(3) At low humidity, 20 to 30 per cent, the temperature of the water after spraying may be 
as much as 8 degrees F. below the temperature of the air and the total drop in temperature 
40 to 45 degrees F. 

(4) The loss of water by evaporation is approximately 0. 15 pound per degree lowering of 
temperature per 100 pounds of water discharged, or a gross loss of about 6 per cent for 40 degrees 
F. lowering of temperature. In no case was the loss found to exceed 7 per cent. 



TABLE 5 



SCHUETTE-KOERTING NOZZLE CAPACITIES 





Siae of Noal« 
in Inches 


Capacitibs in GaujOns pbb Minutb at 


VaBIOUS PaBBBUBBB 




5 Lb. 


6 Lb. 


7 Lb. 


8 Lb. 


9 Lb. 


10 Lb. 


2 


54 

77 

115 


60 

86 

125 


66.5 
92 
133 


70.6 

98 

140 


75 
103 
146 


78 


2H 


106 


8 


151 









Under ordinary atmospheric conditions (air at 70*^ F., and 60 per cent rdative humiditv) the operation of a eon* 
denang and reeooung out^t of this type will be approximatdy as foUows: 



Vacuum 
Inches 


Lb. of 

Water 

Per Lb. 

Steam 


Cofi'c- 

Temp, of 

Cooling 

Water 


Approz. 
Temp, of 
Discharge 

from 

Condenser 

P. 


Degrees 
Reduction 


Temp, of 
Water 
After 

Spraying 


Degrees 
Above 


Reduction 
Obtained by 


28" 
27" 
26" 


50 
35 
30 


20 
29 
34 


92 
105 
110 


20 
29 
34 


72 
76 
81 


2 

6 

11 


Single Spraying 
Single Spraying 
Single Sprayioc 



COOLING PONDS AND TOWERS 



455 



TABLE 6 

LOG OP SPRAY-COOUNG POND 5000-KW. STEAM TURBINE PLANT IN NEW ENGLAND 



OpermtlBg PreMxre, 11 lb. per aq. in. 

Ti -> temperature of dieeharfe water» in degreee F. 
Tt "- t e mperature of water after q>raying, m degreea 
Tt — temperature of surrounding air» in degreee P. 

Hum. — relative htaniditiea taken at 8 P. M . 



8upp(y 



3^(y'Eanh9n Emcan-ment 






eo'o- 






./o'-j*-/o'H 



.'^^S 



11 




i2^Spray Nozzha 
200 H,P.Coofing Pond 



PLAN 




Gr, 




Section Walt 



op9 2fiQn l' 
Grade 



Bottom Won Puddlod 
SECTIONAL ELEVATION ON LINE XX 

Fio. 8. OoouNO Pond wttb Sprat Nozslbb. 



Month 


Humidity 


Temp's 


S AM, 


12 11. 


4PJf. 




Jaaitary 


62% 


« 


[Ti 
Tt 

\Tt 


68* 

48* 

8* 


73* 
58* 
14* 


73* 
58* 
20* 


Clear 

1 


February 


88% 


{Ti 
Tt 
T* 


76* 
64* 

29* 


81* 
61* 
88* 


83* 
68* 
85* 


Cloudy 


March 


60% 


(Ti 
Tt 
Tt 


79* 
58* 
80** 


86* 
66* 
50* 


90* 
70* 
48* 


Clear 


April 


66% 


ITi 
Tt 
Tt 


86* 
66* 
56* 


90* 
71* 
68* 


92* 
73* 
68* 


Clear 


May 


.72% 


ITi 
Tt 

[Tt 


89* 
70* 
65* 


94* 

75* 
72* 


97* 

78* 
70* 


Clear 

1 


June 


90% 


{Ti 
Tt 
Tt 


107* 
78* 
67* 


111* 
83* 
68* 


116* 
85* 
68* 


Cloudy 


July 


70% 


[Ti 
Tt 
Tt 


108* 
90* 
90* 


118* 
98* 

98* 


118* 

98* 

102* 


Clear 


August 


84% 


{Ti 
Tt 
Tt 


112* 
88* 
72* 


114* 
89* 
74* 


116* 
90* 
79* 


1 

Cloudy 


November 


70% 


{Ti 

Tt 

[Tt 


89* 
62* 

27* 


90* 
64* 
88* 


88* 
68* 
84* 


Cloudy 



Area of CooUng Pond Equipped with Spray Nozzles. Based on the average weather condi« 
tious prevailing in the Central and Northern States, it is customary to allow about 1 sq. ft. of 
pond surface for every 200 to 250 lb. of water sprayed per hour for plants above 1000 i.hp. 



456 POWER PLANTS AND REFRIGERA^ON 

Smaller plants will require a somewh&t l&rger area due to the fact that it is desintble to 
keep the spray noulea about 20 (e«t from the edge of the pond. Fig. 8 sbowB tbe deaign suit- 
altle for a 200 hp. plant equipped with twelve ^in. Spray Bngiaefring Co.'t ooiilee. The area 
required for large inatallationa may be calculated on a baaia of uaiog 0.2-in. noulea working at 
ten pounds per sq. in. preeaure, givii^ a discharge of 39,000 lb. water per hour. 

COOLING TOWERS 
Tbe condenaing water coming from either steam or ammonia condenaera ia pumped to tbe 
ti)p of a tower, which is usually filled with wooden or tile checkerwoHi or ^vanised ateel wire 
Hnreena. Tbe water in ila passage down through the checkerwork presento a large evaporatiDg 
ijurfaoe to the air flowing upward tiirougb tbe tower, the cooling of the water being efTected 
principally by the evaporaticxi of a small portion of it. In thecny tike action ia similar to that 
of a humidifier. The air will leave tbe top of the tower 90 to 100 per oent saturated and S to 




Fio. B. Trrse or Wood CHacKBRwaaa fob Coouno Towxbs. 

IS degrees lower than the temperature of the entering water, average figures being 95 per cent 
saturation and 10 degrees lower temperature. 

Tbe limit qf cooling efftd is reached when the water has been reduced in temperature to tbe 
wet-bulb temperature of the entering air at which point ev^Mration ceases. Thia is the tem- 
perature of adiabatic saturation for the gjven condition. Commercial inatallationa vary consider- 
ably in the degree to which they approach this limit. 

PubUsbed tests indicate that the actual drop in temperature tA the water passing through the 
tower will be i4>pro]dnwtely 30 to 60 per cent of the maximum possible drop. 
Let ti — temperature of hot water entering top of tower. 

t, M temperature of water leaving base of tower. 
I — wctrbulb temperature of entering air at base of tower. 
li — 1 = maximum possible drop in temperature of tbe water. 
0.40 (li ~CS= drop in temperature that may ordinarily be obtained in oonunveial io- 

' stallations. 

Then (» - (i - 0.40 (d - ()■ 
B = efficiency of tower. 



COOLING PONDS AND TOWERS 

- 0.30 to 0.50, average value 0.40. 
Qi — heat content o( entering air above 0° F. 
Qt ~ heat content of leaving air above 0° F. 
W — weight of water Ui be cooled per min. 

u - weight of air to be eiroiilated per rein. 

d -^ density of air corresponding to dry-bulb temperature of entering air. 

C = cu. (t. of air measured at diy-bulb temperature of entering air. 



The values of Q, and Qt reay be read direct from the "Paychrometric Chart" (Fig 4), or 
caleulAted by means of the saturated air tables. See Table 7 for examples. 

The air is circulated through cooling towers either by natural draft or by means of fans. 
Fan draft towers, as ordinarily constructed, have an overall height of approximately 30 to 35 
feet, the water being raised to a height of about 28 to 32 feet to the distributing trough. With 
natural draft towers a chimney of approxiniately 40 feet in height is added, making the overall 
height about 75 feet. The water is elevated to the same height as with the fan draft type of 
tower. 



Pia. 10, Pan DSArr Town — Wood ComraocnTON. 

Cooling Tower TesL The following gives the results of a test made on a Wheeler fan draft 
cooling tower plant at Eliubethport, N. J. The tower is workmg in connection with a Whtrler 



458 



POWER PLANTS AND REFRIGERATION 



TABLE 7 

COOUNG EFFECT FOR VARIOUS TOWER EFFICIENCIES 



Wet Bulb 

Tempcfrature of 

Air 

t 



86 
40 
45 
60 
55 
60 
66 
70 
76 
80 
86 



70 
66 
60 
55 
60 
45 
40 
35 
30 
25 
20 



Initial Tbupbraturb of Water ENmuNO Towbb, tu 



106« 



£«0.40 




28 
26 
24 
22 
20 
18 
16 
14 
12 
10 
8 




77 
79 
81 
83 
86 
87 
89 
91 
93 
96 
97 



E=o.eo 




42 
89 
36 
88 
80 
27 
24 
21 
18 
16 
12 




68 
66 
69 
72 
76 
78 
81 
84 
87 
90 
93 



110* 




76 
70 
65 
60 
55 
60 
46 
40 
86 
30 
25 



£7^0.40 




^Qd 



30 
28 
26 
24 
22 
20 
18 
16 
14 
12 
10 




80 
82 
84 
86 
88 
90 
92 
94 
96 
98 
100 



£=0.60 




45 
42 
39 
86 
88 
30 
27 
24 
21 
18 
16 




66 
68 
71 
74 
77 
80 
83 
86 
89 
92 
96 



115« 



80 
76 
70 
65 
60 
66 
60 
46 
40 
85 
80 



£7-0.40 * E^0.€O 



Ofi 


1 


|3 


Jl 


y»^ 




"3*^ 


^d 




li 


<o 


Ch 


82 


88 


80 


86 


28 


87 


26 


89 


24 


91 


22 


98 


20 


96 


18 


97 


16 


99 


14 


101 


12 


108 




•at 



48 
46 
42 
89 
86 



80 
27 
24 

21 
18 



67 
70 
73 
76 
79 
82 
86 
88 
91 
94 
97 



TABLE 8 

RESULTS OF TESTS ON A WHEELER FORCED DRAFT COOLING TOWER 



CooLiNQ Water 








Air 


t 

1 


Cu. Ft. op Air per 
Minute 




Gallons 
Musuta 


Temperature 


Entering 


Outgoing 


EA- 
dency 


(i 


h 


Temp. 


Hum. 


Wet 

Bulb 

Tempt 


Temp. 


Hum. 


Ajiemom~ 
eter 


Cal- 
culatedt 


fit 


651 


105 

107.8 

112 

108.6 

109.9 

116 

135 


84.7 
87.6 
88.6 
87 

90.5 
98 
115.8 


71 
72 
66 

1 69 
83 
43 
60 

i 


40 
60 
60 
48 
48 
75 
73 


57 
63 
56 
57 
69 
40 
56 


90 
93 
96 
92 
96 
101 
118 


100 
100 
100 
100 
100 
100 
100 


63,900 
50,100 
51,400 
50,200 
60,600 
23,600 
17,576 


42,000 

• • • • • 

41,100 


0.42 


638 

638 


0.46 
0.42 


643 

640 

682* 

630» 


0.89 
0.47 
0.23 
0.24 


* In these tests the i 


fan was not 


runninff — ' 


natural ( 


draft. 















t Effideney as calculated by authors ( C — ' _ * j 
X Calculated by authors. 



surface condenser of 280 square feet of cooling surface, mounted over a 10 x 12 x 12 combined 
air and circulating pump. The efficiency (E) has been added by the authors: 

ObservatioDS made on June 24, 1904: 

Temperature of air 81 

Wet bulb, t 60 

Temperature of air at top of tower 80 

Temperature of water in troughs, ti 105 

Temperature of water in tank, it 83 

Revolutions of fan, 239 r.p.m., air pressure. . . ^ inch water. 

Velocity of air out of tower 822 feet per minute. 

Gallons of water per minute passing over mats 385 per minute. 

Vacuum 26 inches. 



n 



COOUNG PONDS AND TOWERS 



459 



Tempeimturo of air-pump discharge 87 degrees. 

Efficiency of tower, B 0.61. 

Obserratioiia made June 28, 1904, 9 A.M. 

Temperature of air 76 degrees. 

Wet bulb, / 59 degrees. 

Temperature of air at top of tower 81 degrees. 

Temperature of wator in troughs, ti 96 degrees. 

Temperature of wator in tank, h 78 degrees. 

RevolutionB of fan, 232 r.p.m., air pressure ?^ inch water. 

Velocity of air out of tower 680 feet per jainute 

Gallons of water passing over mats 406 per minute. 

Vav jum 26.6 inches. 

Temperature of air-pump discharge 90 

Efficiency of tower, E 0.49. 

Obeervations made June 28, 1904, 3 p.m.: 

Temperature of air 74 dcKA«7C9< 

Wet bulb, t 57 degrees. 

Temperature of air at top of tower 83 degrees. 

Temperature of water in troughs, <i 99 degrees . 

Temperature of water in tank, tt 80 degrees. 

Revolutions of fan, 237 r.p.m., air pressure ^ inch water. 

Velocity of air out of tower 769feet per minute. 

Gallons of water passing over mats 470 per minute. 

Vacuum 25.5 inches. 

Temperature of air-pump discharge 92 degrees. 

Efficiency of tower, E 0.45. 

OlMiervations made June 29, 1904: 

Temperature of air 78 degrees. 

Wet bulb, t . . . 71 degrees. 

Temperature of air at top of tower 86 degrees. 

Temperature of water in troughs, ti 108 degrees. 

Temperature of water in tank, tt 82 degrees. 

Revolutions of fan, 241 r.p.m., air pressure ^i inch. 

Velocity of air out of tower 772 feet per minute. 

Gallons of water passing over mats 430 per minute. 

Vacuum 25.5 inches. 

Temperature of air-pump discharge 93 degrees. 

Efficiency of tower, E ' 0.71. 



TABLE 9 
TEST OP WHEELER-BALCKE NATURAL DRAFT COOLING TOWER AT BRISTOL, CONN. 





Test 
Number 


Aug. 
1912 




Watbr 






An 






G. P. M. 


Temp. 
In 


Temp. 
Out 


Temp. 
In 


Temp. 
Out 


Humidity 
In 


Humidity 
Out 


1 


18 
15 
16 
16 
16 
26 
27 


850 
877 
880 
1065 
1068 
850 
850 


114 
104 
94 
92 
100 
114 
104 


84 
86 
76 
76 
74 
87 
80 


84 
80 
68 
70 
70 
81 
77 


• * 

95 
84 
88 
91 
99 
95 


62 
61 
47 
37 
37 
76 
48 


100 


2 


100 


S 


100 


4 


100 


6 


100 


6. 


100 


7.. 


^ 


100 







460 POWER PLANTS AND REFRIGERATION 

Biample. Required Uie unouDt o! lur to be circulated per minute, aiie of fan and power n- 
luired for a coolins lower to cool the circulatins water for a jet type condenser connected to a 500-kw. 
unit. Assumed WHter rale of unit, 20 lb. per kw.-hour 36 in, vacuum referred to a 30-in. barometer. 
Temperature corresponding to vacuum is 125° F.; with a 10° terminal diBerence the temperatiu^ of 
the water leaving will be 115° F. Assume that the average outside air temperature is 65° F. and the 
relative humidity is 60 per cent for the locality in queaboo. 

Referring to the " Paychroinetric Chart," it ia fouad that the wet bulb temperature oorre- 
npondiag to this condition is 57° F. The raaximum theoretical drop id temperature of tho water 



1 



Pig. 11. Fan Draft Coounq Towbr. Tnjt PiLLimi. ( tTorififn^ion.) 

if cooled down to the limit would l>o 115 ^ 57 ^ 58 degs. Id a properly deoigDed cooling tower 
the actual drop in temperature should be about 40 per ceDt (tower efficiency 0.40) of this amount, 
or 58 X 0.40 - 33 degs. The fioal temperature ol the water leaviDg the base of tower and tho 
initio temperature of the circulating water for the condenser may be safely assumed aa 115—23 
- 92° F., say 90' F.. for the conditions Bpeoilied. 



COOLING PONDS AND TOWEI^ 401 

The air leaving top of tower will be assumed 10 degs. lower than the entering water or 115 — 10 
» 105** F. and 95 per cent saturated. 

The heat removed per pound of air circulated will be the difference between the heat content per lb. 
of the air leaving the top of tower and the heat content per lb. of the entering outside air meaeured 
above 0® F. in each case. 

The ** heat content " of a mixture of air and vapor for any condition is given by the formula: 
Q =» Cpa t + xrW in which Cpa = 0.24 sp. ht. of air at constant pressiu-e, t « dry bulb temperature, 
X as relative humidity expressed as a decimal, r « latent heat corresponding to temperature t, W ^ 
wei^t of vapor mixed with 1 lb. of dry air when "saturated *' (100 per cent relative humidity) at tem- 
perature L Heat content of entering air (initial condition) , 65° F. and a relative humidity of 50 per cent. 

Qi = 0.24 X 66 + 0.50 X 1055.5 X 0.0132 = 22.5 B.tu. 

Heat content of leaving air (final condition) 105** F. and a relative humidity of 95 per cent. 

Qi " 0.24 X 105 + 0.95 X 1033.9 X 0.0500 =» 74.3 B.t.u. 

The heat removed per lb. of air circulated is therefore Qt —Qi = 74.3 — 22.5 » 51.8 B.tu. 
The total heat to be removed from the circulating water on a basis of 38 lb. water per lb. of steam 
condensed, corresponding to a 26-in. vacuum, will be: 

500X20 X 39 X<115-90) 

= 162,500 B.t.u. per mm. 

60 

The weight of air to be handled by the fans per min. is therefore 162,500 / 51.8 « 3138 lb. The 
density of ak at 65** F. is approximately 0.0756 lb. per cu. ft. The capacity of fan required is 3138 / 
0.0756 *B 41,500 cu. ft. per min. The total resistance against which the fan is to operate should not 
ordinarily exceed H" water. 

Referring to Table 10 we find that the nearest size disc type fan for the above capacity and pres- 
sure is a 96-in. diam. wheel. The efficiency of this type of fan is approximately 0.33. The brake 

I. / *u r • *u Au 41.500X5^X5.2 ^^ 
horsepower for the fan is then: d.hp. = = 7.4. 

0.33 X 33,000 

Size of Tower and Evaporatiiig Surface. In planning a cooling tower the water should be 
kept in contact with the evaporating surface (checkerwork, mats, etc.) and not allowed to fall 
free. 

The inside area of the to^'er may be approximated by allowing an air velocity of approxi- 
mately 700 ft. per min. through the free area. The area of the evaporating surface may be 
calculated on a basis of 200 B.t.u. per sq. ft. per hour for a 10-deg. drop in the temperature 
of the circulating water and about 700 B.t.u. for a SS-deg. drop. 

Example. The net or free area of tower required for the amount of air given by the preceding 
example is 41,1 10 -t- 700 *= 60sq. ft. The total area will depend upon the type of evaporating surface 
employed. In this example a checkerwork of 1" x 4" cypress boards placed on edge and 5" centers 
will be assumed, the free area being equal to 64 per cent of the total or gross area. The total area 
required is therefore 60 -J- 0.64 = 94 sq. ft. 

For a 25 deg. temperature drop approximately 500 B.t.u. per sq. ft. per hour will be dissipated. 
The total area of evaporating surface required is: 60 X 162,500 -r 500 «= 19,500 sq. ft. 

With the arrangement of evaporating surface stated there will be nearly 8 sq. ft. of surface per 
cu. ft. of checkerwork, then 19,500 -r 8 = 2,438 cu. ft. is necessary. 

This volume is secured by making the checkerwork 10' x 10' x 24' high. The total height of , 
the tower, allowing for the 8 ft. dia. fan and 2 ft. for the distributing troughs, etc., is 34 ft. The catch 
basin or sump at the base of tower may be made about 4 ft. deep and constructed of concrete if set in 
the ground 

Power Reqtiired to Operate Fan Draft Cooling Towers. Assuming a centrifugal pump 



462 POWER PLANTS AND BEPRIGERATION 

efficiency of 0.60 tmd a head of 45 ft. to allow for pipe friction, the brake honepower required 
to pump the oooliog water in the preceding example ia: 



0,60 X 60 X 33,000 



The power required for the fan, previously calculated, is 7.4 d.hp. The total power re- 
qmred wiU be: 14.8 + 7.4 - 22.2 d.hp. 

If pump and fan are each driven by a oiotor having an efficiency of 0.85, the electrical bone- 
power input will be 22.2 -^0.85 = 26.1m- 26.1 
^ 1.34 - Iff .5 kw. This amounts to 10.5 ^ 
SCO or 3.0 per cent of the power Bsnerated by 
the main unit. 

If the fan and pump are each driven by a 
small high-speed steam engine, the water rat« 
of which is 40 lb. per i.hp.-hour and with an 
assumed mechanical efficiency of 87 per cent, 

22.2 X 40 
thesteam required will be — — — — = 1,020 lb. 

per hour. 

Specification and Guarantee. Fur every 
cooling tower a clear ami prccitie guarantee 
fully protecting the intercsls of the purchaacr 
should be pven, embracing. 

Efficiency — temperature from and to which 
the water is to be cooled under given atmo- 
spheric conditions {wet bulb temperature). 

Capacity — amount of water to be cooled. 
I,, Power required for operating the fane 

(maximum and average). 

Durability (according to practical expc- 

*' Workmanship. 

For the computation of correct estimates, 
as well as for the comparison of quotations with 
bids of various manufacturers, full information 
is necessary in regard to all the elements which 
may influence the construction of the cooling 
■"" tower as to si»e, efficiraicy, etc. — vii.; 

1. Type of Cooling Tower and material 
to be used for shell or frame— wood, masonry, 
reinforced conoete, or steel. 
Fio. 18. NATvaAL Draft Cooling Toweb, 2. Location and space available. 

SnowtNG Zioua Coouno Suhfacb. 3. Altitude and atmospheric coniiilioo!! 

prevailing at place of erection. 
4. Amount of water to be cooled per hour — or nuwimum pounds of steam to be condensed 
— if for ice plant, refrigerating capacity per ton. 
6. Temperature of initial water. 

6. Lowest temperature of water required. 

7. Amount of B.t.u. t« be absorbed between temperature range of from to 

In addition the type and construction of the steam (surface or jet) and of the a 



COOLING PONDS AND TOWERS 



463 



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•H CO 10 b-' oi 09 to O oi Ok f-« ^' Ok to 10 
«-if-i090900lOdt*O»^ 



5^1 

00 « 



s s; S S 9 $ 3 3 ^ s s s s s ^ 



464 



POWER PLANTS AND REFRIGERATION 



(atmospheric or double pipe) oondensere, whether gas enginesi reciprocal steam engines or tur- 
binesi should be stated. 

TABLE 11 

APPROXIMATE GROUND AREA FOR MITCHELL-TAPPEN CX)OUNG TOWERS 

(Atmospheric Towers. Average Height 30 Feet) 



GaUoQs per Hour 



1.500 
3.000 
4.500 
6.000 



DimenaioQs, Fee* 




GaUons per Hour 



12.000 
18,000 
24.000 
80.000 



EMmeniiona, Ft 



19 X 19 
19x24.7 
19 X 80.5 
19x36.3 



TABLE 12 
APPROXIMATE DATA FOR FAN DRAFT TOWERS* 



CoouNG Capacity 






H<W8epower 


Siaeand 

No. of 

Fans 

Feet 






Gallons per Hour 


Height 
Feet 


Area of 

Tower 

Base 


Served 
Comp. 
Cond. 


Average 
R.P.M. 


Avera« 
FanHp. 






Ammonia 


Steam 






Eng. 






2.100 


4.200 


26 


19 X19.6 


50 


1-6 


110 


1.26 


8.100 


6.200 


26 


19.8x20.0 


76 


1-6 


160 


1.75 


4.200 


8.400 


26 


20 X20.8 


100 


1-7 


146 


2.26 


6.260 


12.600 


26 


21.6x22.6 


160 


1-8 


146 


3.50 


8.800 


16.700 


26 


28.8x24.6 


200 


1-9 


136 


5.60 


11.600 


21.000 


26 


24.6x25.8 


260 


1-10 


136 


8.00 


12.600 


26.000 


26 


26.6x27.0 


800 


1-10 


146 


11.00 


17.100 


84.200 


27.6 


27.6x29.6 


400 


1-12 


116 


14.00 


20.760 


41.500 
_» 


27.6 


29 x80.0 


500 


1-12 


145 


18.00 



*ract& 



* "PracMbal Engineer.'* January. 1916. 

Cost of Cooling Tourers. On a basis of a 26-in. vacuum referred to a 30-in. barometer, cool- 
ing tower costs, erected in place, are approximately 6 to 7 dollars per kw. rating of direct-connected 
units. 

The following figures are actual costs of towers f.o.b. factory, exclusive of the motors or 
engines to drive the fans, rated on a basis of cooling water from 110^ to 80^, vacuum 26", ratio 
of cooling water to steam 1 : 50, water rate of unit 22.5 lb. per kw.-hour. 



TABLE 13 



Kw. Rating 
of Unit 


GaUons 
Water 

Mmute 


Pound 
Water 

Mtoute 


Siae 
Tower 


* 

<»•» Total 
^, Weight 
*^°" Pounds 


Pries 
F.03. 
Faetory 


400 


900 
2.260 


7,500 
18.750 


10'xll'x40'ht. 
14'xl6'x42' ht. 


2- 7' dia. isonn 


$2,700 
6.000 


1,000 


2-10' dia. 


70.000 





r •• 



CHAPTER XV 



PIPE, FITTINGS, VALVES, COVERINGS AND ACCESSORIES 

PIPE 

Conundfeitl Clataiflcttion of Pipe. Commercial pipe is made of wrougkt-iron or mild sUel, 
ill certain definite sizes, always stated in terms of the nominal internal diameters up to and in- 
cluding 12". (Table 1.) Above 12" internal diameter the size is based on the outside diameter, 
and the thickness of metal always specified. 

There are three weights or strengths of pipe generally recognized in engineering practice', 
known as "standard,'* "extra strong*' and "double extra strong," all of which have the same 
outside diameter for a given size. * 

Standard Pipe. Standard pipe is also known as JnU weight pipe and is made from sheets 
<^ sufficient thickness to permit of the necessary manipulation, such as heating and rolling, and 
still finish in random lengths of from 18 to 20 ft. which will weigh, including coupling on one 
end, within 5 per cent of "card weight" (Table 1). Unless otherwise specified, this pipe is 
furnished in random lengths with threads and couplings. 

TABLE 1 

DIMENSIONS OP STANDARD AND EXTRA STRONG* WROUGHT-IRON AND STEEL PIPE 







DlAMETBR 




CiRCUIfFBRBNCB 


Internal 


Length 
Pipe in 


Nominal Weight 
















Transvbrsb Arba 


Ub. pe 


r Foot 




























N<munal 


Extnnml 


Internal 


External 


. Internal 






Ft. per 






Siae 


Standard 






SUndard 










Square 








and 
Extra 
Strong 






and 
Extra 
Strong 






SUnd*ard 


Extra 
Strong 


Ft. of 
Exter'l 
Surface 


Standard 


Extra 
Strong 




Standard 


Extra 
Strong 


Standard 


Extra 
Strong 


H... 


0.405 


0.269 


0.215 


1.272 


0.848 


0.675 


0.0673 


0.0363 


9.440 


0.244 


0.314 


H . . • 


.640 


.364 


.802 


1.696 


1.144 


.949 


.1041 


.0716 


7.075 


.424 


.535 


»^»... 


.675 


.493 


.423 


2.121 


1.552 


1.329 


.1917 


.1405 


5.657 


.567 


.738 


H... 


.840 


.622 


.646 


2.639 


1.957 


1.715 


.3048 


.2341 


4.647 


.850 


1.087 


»4... 


1.050 


.824 


.742 


3.299 


2.589 


2.331 


.5333 


.4324 


3.637 


1.130 


1.473 


1 ... 


1.315 


1.049 


.957 


4.131 


3.292 


3.007 


.8626 


.7193 


2.904 


1.678 


2.171 


IK... 


1.660 


1.380 


1.278 


5.215 


4.335 


4.015 


1.496 


1.287 


2.301 


2.272 


2.996 


1'2... 


1 900 


1.610 


1.500 


5.969 


5.061 


4.712 


2.a38 


1.767 


2.010 


2.717 


8.631 


2 ... 


2.375 


2.067 


1.939 


7.461 


6.494 


6.092 


3.366 


2.953 


1.608 


8.652 


5.022 


2'a... 


2.875 


2.469 


2.323 


9.032 


7.763 


7.298 


4.784 


4.238 


1.328 


6.793 


7.661 


3 ... 


3.600 


8.068 


2.900 


10.996 


9.636 


9.111 


7.388 


6.605 


1.091 


7.675 


10.252 


:)».i... 


4.000 


8.548 


3.364 


12.566 


11.146 


10.568 


9.887 


8.888 


0.955 


9.109 


12.505 


4 ... 


4.600 


4.026 


3.826 


14.137 


12.648 


12.020 


12.730 


11.497 


.849 


10.790 


14.983 


4>i... 


6.000 


4.606 


4.290 


15.708 


14.162 


13.477 


16.961 


14.454 


.764 


12.538 


17.611 


5 ... 


6.563 


6.047 


4.813 


17.477 


15.849 


15.121 


19.990 


18.194 


.687 


14.617 


20.778 


... 


6.625 


6.065 


5.761 


20.813 


19.054 


18.099 


28.888 


26.067 


.577 


18.974 


28.573 


i ... 


7.625 


7.028 


6.625 


23.955 


22.063 


20.813 


38.738 


34.472 


.501 


23.544 


38.048 


8 ... 


8.625 


7.981 


7.625 


27.096 


26.076 


23.956 


50.040 


45.664 


.443 


28.644 


48.388 


9 ... 


9.625 


8.941 


8.625 


30.238 


28.089 


27.096 


62.776 


58.426 


.397 


33.907 


48.728 


10 ... 


. 10.750 


10.020 


9.750 


33.772 


31 .477 


30.631 


78.839 


74.662 


.356 


40.483 


64.735 


11 ... 


. 11.750 


11.000 


10.750 


36.914 


34.658 


33.772 


95.033 


90.763 


.325 


46.557 


60.075 


12 ... 


12.750 


12.000 


11.750 


40.055 


37.700 


36. 9U 


113.098 


108.43 


.299 


49.662 


65.415 



NoTB. — Dimensions are nominal and, except where noted, are in inches, 
* Often called extra heavy pipe. 

465 



1 



466 



POWER PLANTS AND REFEUGERATION 



A lighter weight of standard pipe, in sizes up to 6'', known as merchtmi pipe, and ninning about 
10 per cent below ''card weight/' has been discontinued by the principal manufacturen. UnleflB 
this pipe is wanted, it is necessary to specify ''full weight'' pipe. 

Extra Strong Pipe. Extra strong pipe (Table 1) is usually specified for steam, gas or hydraulic 
work at pressures above 125 lb. gage. This pipe is made in randcHn lengths of from 12 to 20 ft. 
and is always furnished with plain ends unless otherwise specified, although as Inuch as 10 per 
cent of a total order may be in lengths from 6 to 12 ft. 

Double extra strong pipe is omitted from Table 1 since its use is limited almost entirely to 
high-pressure hydraulic work. The same trade practice is followed in furnishing it as for extm 
strong pipe. 

Outside Diameter Pipe. Outside diameter pipe, known as O. D. pipe (Table la), is the com- 
mercial designation appUed to all regular sizes above 12". Since the terms standard or extra 
strong do not apply to these sizes, it is always necessary to give the thickness as well as the outside 



TABLE la 

OUTSIDE DIAMETER (O. D.) STEEL PIPE 
Nominal weight in pounds per foot 



Sise 
Outaide 










Thickness • 


























Diam. 


Mln. 


Vii In. 


% In. 


Vi« In. 


HIn. 


•/«In. 


HIn. 


"/it In. 


kin. 


14 


36.75 


45.72 


54.61 


63.42 


72.16 


80.80 


89.86 


97.84 


106.20 


15 


89.42 


49.06 


58.62 


68.10 


77.50 


86.81 


96.08 


106.20 


114.20 


16 


42.09 


52.40 


62.63 


72.78 


82.85 


92.83 


102.70 


112.60 


122.20 


17 


44.76 


65.74 


66.64 


77.46 


88.19 


98.84 


109.40 


119.90 


180.30 


18 


47.44 


59.08 


70.65 


82.14 


93.54 


104.80 


116.10 


127.20 


188. SO 


20 


62.78 


65.76 


78.67 


91.49 


104.20 


116.90 


129.40 


141.90 


154.30 


21 


65.45 


69.10 


82.68 


96.17 


109.60 


122.90 


136.10 


149.30 


162.80 


22 




72.44 


86.68 


100.80 


114.90 


128.90 


142.80 


166.60 


170.30 


24 




79.18 


94.70 


110.20 


125.60 


140 90 


166.20 


171.30 


186.80 


26 


• • • • • 




102.70 


119.60 


186.30 


152.90 


169.60 


186.00 


202.40 


28 






110.70 


128.90 


147.00 


165.00 


182.90 


200.70 


218.40 


80 








188.20 


157.70 


177.00 


196.30 


216.40 


234.40 



diameter. This pipe is furnished in random lengths of from 8 to 20 ft., depending on the sise, 
and with plain ends. The threading of O. D. pipe is not recommended. 

In connection with pipe sizes, Table 2, giving certain tube data, may be found to be of service. 



TABLE 2 

TUBE DATA, STANDARD OPEN-HEARTH OR LAP-WELDED STEEL TUBES 



Sice 

Extern. 

Diam. 



14-. 
\M.. 

14.. 
2 .. 
2 .. 
2 .. 
3k.. 
3K.. 
3^.. 
4 .. 
4 .. 
4 .. 



B.W. 


Thick- 


Intomal 


Gage 


nesB 


Diam. 


10 


.134 


1.232 


9 


.148 


1.204 , 


8 


.166 


1.170 


10 


.134 


1.782 


9 


.148 


1.704 
1.670 


8 


.165 


11 


.120 


3.010 


10 


.134 


2.982 


9 


.148 


2.954 


10 


.134 


3.732 


9 


.148 


3.704 


8 


.165 


3.670 



CnCUMFBRKNCB 



External 



4.712 

4.712 

4.712 

6.283 

6.283 

6.283 

10.210 

10.210 

10.210 

12.666 

12.666 

12.666 



Internal 



3.870 

3.782 

8.676 

6.441 

5.353 

5.246 

9.456 

9.368 

9.280 

11.724 

11.636 

11.530 



Transvebsb Arba 
Square Ihchib 



External 



1.7671 
1 .7671 
1.7671 
3.1416 
3.1416 
3.1416 
8.2958 
8.2968 
8.2968 
12.666 
12.666 
12.566 



Interaal 



1.1921 
1.1386 
1.0751 
2.3660 
2.2778 
2.1904 
7.1157 
6.9840 
6.8636 
10.989 
10.776 
10.678 



Square 
F^of 

External 
Surfaee 
per Ft. 

of Length 



.892 

.392 

.892 

.623 

.523 

.623 

.850 

.860 

.850 

1.047 

1.047 

1.047 



Length 
in fwt 
per Sq. 
Foot of 
External 
Surfaee 



2.646 

2.546 

2.546 

1.909 

1.909 

1.909 

1.176 

1.176 

1.176 

.964 

.954 

.964 



NoniBal 
Weight 
Pounds 
per Ft. 



1.966 
2.187 
2.863 
2.670 
2.927 
3.284 
4.011 
4.469 
4.908 
6.682 
6.000 
6.768 



Note. — Dimensions are nominal and, except where noted, are in inches. 



PIPE, FITTINGS, VALVES, CX)VERING8 AND ACCESSORIES 



467 



Threading Pipe. The threading of either wrought-iron or steel pipe requires suitable dies 
tftdapted to the metal to be cut. Dies suitable for wrought iron will tear steel pipe, and hence 
the complaint is sometimes made that steel pipe is brittle. This can be readily overcome by using 
proper dies. All pipe is threaded uniformly using Briggs standard gage and taper. This taper 
of ^" to I'-O" on all standard pipe threads is necessary in order to secure a tight joint in the 
threads when screwing the pipe into a fitting or valve. 

Testing P^. Pressure tests at the mill of wrought-iron or steel pipe are commonly made in 
order to show the presence of flaws or other defects in the wM or body of the pipe. Wrought 
pipe, as distinguished from seamless tubing, is either hiUi or lajHvelded; sizes up to and including 
l}4" being made by the former, and those 1^" and larger by the latter process. Lap-Welded 
pipe, IH" diameter, may safely be tested to 2500 lb. per sq. in. cold hydraulic pressure, while 
12" diameter pipe should not be tested to more than 300 lb. per sq. in. The makers vary the 
test pressure in accordance with the diameter so as to produce approximately the same fiber 
stress in each size of pipe. 

The thecreticcd bursting pressures for steel pipe of varying diameters ranging from 14" 
diameter to 12" diameter can be calculated, and are gpven by John B, Berryman in Table 3. 

TABLE 3 

THEORETICAL BURSTING PRESSURE OF WROUGHT-IRON PIPE 
Baaed on New Material with Plain Enda. Wel