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The following have contributed manuscript or data or have otherwise assisted 
in the preparation of this work: 

Edmoxd Siroky 
I. Y. Le Bow E. R. Powell, 

Terrell Croft Engineering Company 


By Terrell Croft 

American Electricians' Handbook 

Wiring of Finished Buildings 

Wiring for Light and Power 

Electrical Machinery 

Practical Electric Illumination 

Practical Electricity 

Central Stations 

Lighting Circuits and Switches 

Alt:^rnating-Current Armature Winding 

Conduit Wiring 

Electrical Machinery and Control 

Circuit Troubles and Testing 
Electrical-Machinery Erection 
Signal Wiring 
Automobile Wiring Diagrams 


Terrell Croft 


Steam Boilers 

Steam-Power- Plant Auxiliaries 

Steam-engine Principles and Practice 

Steam-turbine Principles and Practice 

Machinery Foundations and Erection 

Practical Heat 












Ss^y Division 

First Edition 
Fourth Impression 


LONDON: 6 & 8 BOUVERIE ST., E. C. 4 



Copyright, 1923,/ by Terrell Croft 





Although the steam turbine is a relatively new development 
in steam power-plant practice, it is already of great importance. 
Its adoption has, because of its economic superiority for many 
conditions, been very rapid. Today, turbines of different 
capacities ranging from 1 hp. up to 80,000 hp. are being effec- 
tively utilized for power generation. The number of turbines 
in use will soon exceed — if it does not already exceed — the 
number of reciprocating steam engines. It follows that all 
successful power-plant men must now be informed concerning 
these machines. 

Steam-tuebine Principles and Practice has been pre- 
pared, for the ''practical" man, to furnish this information. 
It has been written to provide the operating engineer, the 
plant superintendent, or manager with such steam-turbine 
information as he requires in his everyday work. The aim 
has been to treat only topics of two general classes: (1) Those 
with which a man must he familiar to insure the successful and 
economical operation of steam turbines. (2) Those a knowledge 
of which is necessary to enable a man — one who is not familiar 
with the details of its design or theory — to make a wise choice if 
he contemplates the purchase of a turbine. Only sufficient 
theory is given to insure a sound understanding of the principles 
of turbine operation. The ''design" of turbines is not treated 
at all. A working knowledge of arithmetic will enable one to 
read the book intelligently. 

Drawings for nearly all of the 282 illustrations were made 
especially for this work. It has been the endeavor to so design 
and render these pictures that they will convey the desired 
information with a minimum of supplementary discussion. 

Throughout the text, principles which are presented are 
explained with descriptive expositions or worked-out arith- 
metical examples. At the end of each of the 14 divisions there 
are questions to be answered and, where justified, problems to 



be solved by the reader. These questions and problems are 
based on the text matter in the division which they follow. 
If the reader can answer the questions and solve the problems, 
he then must be conversant with the subject matter of the 
division. Detail solutions to all of the problems are printed 
in the appendix in the back of the book. 

As to the method of treatment: Fundamental principles of 
turbine operation are first presented so as to provide a knowl- 
edge of the theory which is necessary for the understanding of 
how turbines function. This is followed by a division on 
turbine classifications and nomenclature. Next, the principal 
turbine types and constructions are described and discussed. 
Then follows a division on steam-turbine installation. This is 
followed by four divisions on important turbine parts which 
require periodic attention. These divisions treat of: Shafts, 
bearings, and packing glands; governors and valves; reduction 
gears and couplings; regenerators and condensers. Next is an 
important division on high-pressure, bleeder, mixed-pressure, 
and exhaust-steam turbines. 

Following this are practically-treated divisions on lubrica- 
tion and operation and maintenance. The next division on 
testing explains the methods whereby the efficiencies of tur- 
bines are determined. The last two divisions — one on the 
effects of steam pressure, superheat, and vacuum on steam- 
turbine economy; the other on steam-turbine economics and 
selection — outline the processes by which the most economical 
steam conditions and the most economical turbine may be 
selected for a given installation. They also interpret the effects 
of steam-condition changes on the economy of the turbine. 

With this, as with other books which have been prepared by 
the editor, it is the sincere desire to render it of maximum use- 
fulness to the reader. It is the intention to improve the book 
each time it is revised and to enlarge it as conditions may 
demand. If these things are to be accomplished most effec- 
tively, it is essential that the readers cooperate with us. This 
they may do by advising the editor of any alterations which 
they feel it would be advisable to make. Future revisions and 
additions will, insofar as is feasible, be based on such 
suggestions and criticisms from the readers. 


Although the proofs have been read and checked very care- 
fully, it is possible that some undiscovered errors may remain. 
Readers will confer a favor in advising the editor of any such. 

Terrell Croft. 
University City, 
St. Louis, Mo,, 
December, 1922. 


The editor desires to acknowledge the assistance which has 
been rendered by a number of concerns and individuals in the 
preparation of this book. 

Portions of the text material appeared originally as articles 
by the editor in certain trade and technical periodicals among 
which are Power, Power Plant Engineering, and Southern 
Engineer. In all such cases and in others where material 
from publications has been used, it is beheved that proper 
acknowledgment has been accorded at the proper place in the 

The list of manufacturers who cooperated in supplying text 
data and illustrations would include practically all manufac- 
turers of steam turbines, both large and small, in the United 
States. In virtually all cases where such data have been used, 
special acknowledgement is accorded in the text. The editor 
is particularly indebted to the Allis-Chalmers Manufacturing 
Company and the Terry Turbine Company for the data which 
was submitted by their chief turbine engineers. 

Assistance and information have been obtained from certain 
recognized technical turbine books and, in some cases, tables 
and other data were taken from them. Proper acknowledg- 
ment is accorded in the text wherever such data were used. 

Special acknowledgment is hereby accorded Edmond 
Siroky, Head Mechanical Engineer of the Terrell Croft Engi- 
neering Company, who has been responsible for the technical 
accuracy of the book. 

Other acknowledgments have been made throughout the 
book. If any has been omitted, it has been through oversight 
and, if brought to the editor's attention, it will be incorporated 
in the next edition. 

Terrell Croft. 



Frontispiece iv 

Preface vii 

Acknowledgments x 

Division 1 — Steam-turbine Fundamental Principles 1 

Division 2 — Steam-turbine Nomenclature and Classification. 23 

Division 3 — Steam-turbine Types and Construction 42 

Division 4 — Steam-turbine Installation 74 

Division 5 — Steam-turbine Shafts, Bearings, and Packing 

Glands 86 

Division 6 — Steam-turbine Governors and Valves 119 

Division 7 — Steam-turbine Reduction Gears and Couplings . 160 
Division 8 — Steam-turbine Regenerators and Condensers. . 173 
Division 9 — High-pressure, Bleeder, Mixed-pressure, and 

Exhaust-steam Turbines 186 

Division 10 — Steam-turbine Lubrication 218 

Division 11 — Steam-turbine Operation and Maintenance . . .235 

Division 12 — Steam-turbine Testing 251 

Division 13 — Effect of Steam Pressure, Superheat, and 

Vacuum on Steam-turbine Economy 278 

Division 14 — Steam-turbine Economics and Selection 307 

Solutions to Problems 337 

Index 339 



Wheel ■■ 



1. A Steam Turbine Operates By Virtue Of Heat which it 
abstracts from the steam and which it converts into mechanical 
work. Heat and mechani- 
cal work are both forms of 
energy (see the author's 
Practical Heat) and can, 
therefore, be converted one 
to the other by the proper 
means. Any apparatus 
which can convert heat 
e n e r g y into mechanical 
work is^called a heat engine^ 
Thus, the steam turbine is 
just as much a heat engine 
as is a steam or internal- 
combustion engine. The 
steam turbine is different, 
howeyer ^inTXher^ anner in 
which itconxerts the heat 
e n e r gy into mechanical 
work, as will be_explained. 


Fig. 1. — The earliest known heat engine, 
described by Hero of Alexandria about 150 
B. C, was a reaction turbine. 

Note. — The Steam Turbine Was The First Form Of Heat Engine 
To Be Developed And The Latest To Be Perfected, thus it might be 
said that it is the oldest as well as the newest form of heat engine. The 
earliest record of any heat engine is in a book written by Hero of Alex- 
andria, probably about 150 B.C., in which a steam reaction wheel (Fig. 
1) is mentioned. The next development on record was the turbine of 



Branca, an Italian architect, (Fig. 2). These turbines will be described 

The first turbine patent in the United States was issued in 1831 and 

under it several turbines were built. 

Reducfhn Gears 


.Turbme Wheel 



Fig. 2. — Branca's impulse turbine (1629). 

They seemed to give satisfaction 
for some time but they did not 
last long. 

The commercially successful tur- 
bine is due, however, to the inven- 
tive genius and experiments of 
De Laval and Parsons, who worked 
separately and along different lines, 
during the years 1882 to 1889. 
Turbines of large capacities have 
been developed only within the 
last twenty years. 

2. In A Steam Turbine, Heat Energy Is First Converted 
Into Velocity Energy Or Kinetic Energy which is then converted 
into mechanical work. The fact that steam will issue with con- 
siderable velocity through any small opening in a containing 
vessel is, no doubt, known to all. It will be shown that the 
velocity is derived from heat energy which the steam hberates 
as it passes through the opening. It will also be shown that 
the velocity of the issuing steam can be forecasted with 
reasonable accuracy. Such calculations are, however, the 
work of the turbine designers and, except in so far as they 
explain fundamental principles of turbine operation, will not 
be treated herein. 

3. A Steam Turbine Does Mechanical Work By Virtue Of 
The Velocity With Which The Steam Strikes Or Leaves 
Moving Parts. — As the steam attains its velocity, by issuing 
from an opening, its velocity energy may be converted into 
mechanical work by suitably deflecting its current. In this 
respect, a steam jet acts just as does a water jet. The manner 
in which fluid jets may produce forces will now be shown. 

4. The Terms "Impulse" and "Reaction'* Have Specific 
Meanings In Turbine -engineering Parlance. — These specific 
meanings, which are employed in this book and which are 
explained in the following sections, are different from the 
meanings of the same words as they are employed in physics, 
mechanics and in ordinary usage. 

Sec. 5] 


Note. — "There Is Little Connection Between The Usual 
Meanings Of The Words 'Impulse' And 'Reaction' And The Spe- 
cific Ideas They Are Intended To Convey In Steam-turbine Par- 
lance. Actually, all commercial steam turbines work by both impulse 
and reaction. A German writer has used instead of * impulse ' and ' reac- 
tion' words meaning 'equal pressure' and 'unequal pressure,' which to 
the author seem much more appropriate." This paragraph abstracted 
from M oyer's Steam Turbines, John Wiley & Sons. 

6. An Impulsive Force Or "Impulse" Is That Force Which 
Is Produced On An Object When A Fluid Jet Strikes The Object, 

Nozz/e Tends To'Reacf" In 
This D/recf ion. Fireman 
Must Forcibly Moid If 
Against This Reaction. 

House -y 

Jet Of Water 

Fig. 3. — Illustrating the "impulse" effect of a jet of water when directed against a 
stationary object (window pane). 

Fig. 3. — This is the specific turbine-engineering definition; 
see Sec. 4. The fact that a stream of water from a fire hose 
can shatter a pane of glass (Fig. 3) or even move heavier 

Blocl< OfWoocf Scale Shows 
Fastened ToScale-: Force ^ - 

Jet Of Water '■' || 

S^^////Z Household Scale 
' Secured To Wall' 

Fig. 4. — Measuring the "impuls 
jet of water. 


of a 

Shaft- . Vaned 

'Steam Pipe Vanes '' |J 

Direction Of Rotation 

Fig. 5. — The elements of Branca's steam 
turbine (1629). 

objects which it strikes is well known. A distinguishing 
characteristic of an impulsive force is that the fluid stream 
which strikes the object, and thereby produces the force, leaves 
the object at the same or at a less velocity than that with 
which it strikes the object. A simple means of measuring 
the impulsive force is shown in Fig. 4. The force which the 




scale indicates is that which is produced by the impulse of 
the water jet. 

Note. — This ''Impulse" Principle Was Employed By Branca In 
His Primitive Turbine, Figs. 2 and 5, which was formed by mounting 
a number of vanes on the circumference of a wheel and directing a steam 
jet against them at one point. As the wheel is moved by the impulsive 
force, the steam jet plays successively on other vanes, thus providing a 
continuous motion of the wheel whereby it may be caused to do work. 
Those turbines which employ the impulse principle mainly for their 
operation are called impulse turbines; see Sec. 30. 

6. A Large Impulsive Force Is Produced When A Fluid 
Stream Strikes An Object And Then Leaves It In An Opposite 

Wocvlen Block Scale Shorn ^ 
Nozzle^ Fastened To Scale\ ^ f-Q^f^g . ^ 

^^^Kjef Of Water. ■■ 

;5^^^// HousehoklScale: 

Secured To Wall Wall- 

Fig. 6. — Measuring the impulse of a jet of water. (Compare with Fig. 4.) 

Direction. — This may well be explained by a simple experi- 
ment, Fig. 6. When the fluid stream strikes an object which 
is so shaped that it reverses the direction of the stream, a 

.• Direction Of Rotation 

Fig. 7. 

^^^. \. Jet Of Water 
-A Pelton watcrwheel. 

much greater impulsive force is produced than when the 
direction of the stream is not reversed (Fig. 4). This occurs 
in spite of the fact that the stream may leave the object with 
the same velocity as that with which it approached the object. 

Sec. 7] 


In fact, it is found that the force in Fig. 6 is just twice that of 
Fig. 4. (In ordinary parlance — not in steam-turbine parlance 
— the force produced by the jet leaving the object would be 
called a ''repulsive" force.) 

Note. — The Principle Of Thus Utilizing Large Impulsive 
Forces Is Employed In The Pelton Waterwheel (Fig. 7) and in the 

Supply Sfeam 

Fig. 8. — The De Laval trade mark 
which illustrates the principle of the 
so-called "impulse" turbine. (It uti- 
lizes impulsive forces.) 

I Assembled Turbine 

ETranversc Section HL Longitudinal 
Through Bucket Section 

Through Bucket 

Fig. 9. — An early form of steam tur- 
bine which was patterned after the Pelton 

De Laval single-stage turbine (Fig. 8). Some of the earlier turbines, as 
shown in Fig. 9, were made very similar to the Pelton waterwheel, but 
exactly this construction is no longer used (see Div. 3) because more 
efficient ones have been developed. 

7. A Reactive Force Or "Reaction" Is That Force Which 
Is Produced On An Object When A Fluid Stream Leaves 
The Object At A Greater Velocity Than That With Which It 
Approaches The Object Figs. 1 and 10. This is the specific 
turbine-engineering definition; see Sec. 4. Examples of this 
force are, no doubt, familiar to everyone even though many 
people do not know their explanation. Some familiar 


examples are: (1) The " push-hack '^ or " kick-back '^ of a hose, 
as in Fig. 3, or of a shot-gun. (2) The fireworks pin wheel, 
Fig. 10-7. (3) The revolving lawn sprinkler, Fig. 10-/7. The 
existence of a force when a fluid stream leaves a body may 
well be illustrated by the simple experiment, shown in Fig. 11 
wherein the bucket is held from the vertical by the reactive 
force of the water jet. In each of these cases the velocity 
of the fluid stream which leaves the body is greater than 
that of the fluid stream approaching the body. 


I-Lawn Sprinkler 

Fig. 10. — Illustrative examples of reac- 
tion wheels. 


I- No Jet 

fbinter Shows Deflect/on 
n.- Jet Producing Reaction 

Fig. 11. — A simple experiment which 
proves the existence of a reaction force 
when a jet leaves an object. (The deflec- 
tion is shown exaggerated for clearness.) 

Note. — Reactive Forces, Which Were Produced At The Ends 
Of The Arms Of Hero's Turbine (Fig. 1) as steam issued from them, 
provided the rotational motion whereby work was done. Hero's turbine 
was, thus, similar to our common lawn sprinkler. Fig. 10-/7. No 
modern turbines employ only reaction forces (see Sec. 31), but those 
turbines which employ the reaction principle mainly for their operation 
are called reaction turbines. 

Explanation. — The Nature Of Reactive Forces can easily be 
understood by a study of Figs. 12, 13, and 14. Imagine a tank which 
has similar holes on opposite sides near its bottom, Fig. 12. If both 
holes are corked and the tank is filled with water, the water will exert a 
force on each cork tending to push it out. But, since the corks plug 

Sec. 7] 


equal holes and since both are subjected to the same pressure, the force 
on one cork is equal to that on the other. Say each cork is subjected to 
2 lb. as in Fig. 12-7. If, now, one cork is removed as in II, then the one 
force of 2 lb. is removed and the other 2-lb. force must be balanced from 

'Thcfionless Surface 
I- No React ive" 1- Reactive Force 

Force On Tank 

Fig. 12. — Illustrating how a reactive 
force comes into action. Note that the 
reactive force would exist even if the tank 
were in a vacuum. 


Jef ■■■■■' 
IT- Rotation Produced 
By Two 5 Lb. Reactive 

Fig. 13. — Showing the nature of 
the forces that cause rotation of a 
lawn sprinkler. 

without the tank as shown. If the scale were not in a position to prevent 
it as shown, the remaining 2-lb. force in Fig. 12-/7 would be capable of 
moving the tank and thus doing work upon it. 



:zir:_-_ ) 

^' ^^^'^"^^ 









I" Turbine Rotor Plugged. No Reaction 

E-Rotor Tends To Rotate 

Fig. 14. — Illustrating the nature of reaction forces in a modern reaction turbine. 
The blades or nozzles are here shown with square corners merely to explain the principle. 
Actually, the nozzles of reaction turbines are formed with smooth curved surfaces as 
shown in other illustrations. 

The mode in which reactive forces cause rotation is shown in Fig. 13. 
In 7 the pipes are capped at all ends and the forces are balanced as shown. 
When, however, the pipes are opened as in 77, certain forces are removed 
and the remaining forces are free to turn the sprinkler. 


In Fig. 14-7 the rotor of a modern reaction turbine is diagrammatically 
shown with its outlet plugged. The internal forces on the plug and on 
the equal area at the inlet end of the nozzle are balanced as indicated. 
If, however, the plug is removed as in 77, the force which it withstood 
is also removed and the force on the rotor blade is unbalanced. Hence, 
this force is capable of rotating the rotor. 

8. Steam Liberates Heat As It Flows Through An Opening 

from a region of high pressure to one of lower pressure. Dry 
saturated steam at low pressures contains less heat (B.t.u.) 
per pound than does dry saturated steam at high pressures; 
a study of any steam table will confirm this statement, see 
also the author's Practical Heat. Therefore, if dry satu- 
rated steam undergoes a reduction in pressure, it must either 
lose heat or become superheated. Experience shows, how- 
ever, that when steam expands in a turbine nozzle the steam 
does not become superheated — in fact, it undergoes a reduc- 
tion in quality or dryness. Therefore, in a turbine, the steam 
loses or liberates heat as its pressure is reduced. Experience 
further shows that steam, when it passes without friction 
through an opening, loses just as much heat as it would have 
lost, had it expanded adiabatically behind a piston from the 
same initial to the same final pressure. But, in flowing through 
a relatively small opening, the steam acquires a high velocity 
which it would not have acquired if it had expanded behind a 
piston. It can be conceived that, in flowing through an open- 
ing, the steam does its work upon the steam immediately ahead 
of it by pushing it forward so hard as to increase its velocity. 

9. The Kinetic Energy Which Steam Acquires in flowing 
through an opening from a region of high pressure to one of 
lower pressure is equal (except for small losses) to the heat 
which is liberated by the steam. It is known that the work 
which steam does when expanding adiabatically behind a 
piston is equal to the heat that is liberated by the steam; see 
Practical. Heat. Hence, it follows, that the kinetic energy 
acquired by steam in flowing through an opening is equal to 
the work which the steam would have done if it had expanded 
adiabatically behind a piston. Obviously then, if a steam 
turbine could utilize all of the kinetic energy which its steam 
acquires, it would do exactly the same amount of work as 


would a perfect steam engine which expanded steam adiabati- 
cally between the same pressures. The relation between heat 
energy and kinetic energy in a steam turbine is, therefore, 
neglecting small losses: 

(1) Kinetic energy acquired = 

Heat liberated by adiabatic expansion. 

10. The Velocity Theoretically Acquired By Steam In Flow- 
ing Through An Opening from a region of high pressure to a 
region of lower pressure may be readily computed if the state 
of the steam at the higher pressure is known and if the lower 
pressure is known. These known factors determine the 
amount of heat liberated by an adiabatic expansion. Hence, 
by For. (1), they also determine the kinetic energy acquired. 
The formula, (see also chart C of Fig. 15) which expresses the 
velocity acquired and which is derived below, is: 

(2) V = 223.7%! -H2 (ft. per sec.) 

Wherein: v = velocity acquired by the steam, in feet per 
second. Hi = total heat of the steam at the higher pressure, 
in British thermal units per pound; this may be obtained from 
a steam table or from the chart of Fig. 15. H2 = total heat 
of the steam after adiabatic expansion to the lower pressure, 
in British thermal units per pound; this may be obtained most 
readily from the chart of Fig. 15 as explained below. 

Derivation. — From the mechanics of moving bodies, the kinetic energy 
of any moving body is : 

(3) Kinetic energy = -^f = "2X32^ = -qU (^^-^b-) 

Wherein : W = the weight of the body, in pounds, v = the velocity of 
the body, in feet per second, g = the acceleration due to gravity, in 
feet per second per second (= 32.2). 

Also, from the thermodynamics, see the author's Pbactical Heat: 

(4) Heat liberated by adiabatic expansion = W(Hi — H2) (B.t.u.) 
Or, since 1 B.t.u. = 77Sft.-lb., expressing the same thing in foot-pounds: 

(5) Heat liberated by adiabatic expansion = 778"W(Hi — H2) (ft.-lb.) 

Wherein: W = the weight of the steam, in pounds. Hi and H2 have 
the meanings given above. 778 = the equivalent of 1 B.t.u. in foot- 



Ti'''n"i' '' i' ' i' ' ' i ' i' i ' i | l ' i'l'i' i ^i ' l i ' i | i |i|i 
P BSBmgPerLhg 

!£; Lb. Per Horse power-Howr 

|ik'i|ji,ii'u'J i iV i l|ii' i' iliMi4 i 'ii|i | i | li i i;ii' | i|^'|lMivX^ 

1.4 Ln+ropy 1.5 


Qudli+y Lines 

Lines Of Constant 

-Pressures Above Atmospheric Line 
Given In Lb. Per Sguotre Inch Ga ge 

Pressures Below Atmospheric Line 
Are Given In Inches Of Mercury 
Referred, To, 30 Inch Barometerf 

I I I I I I 

I I I I 

1.6 1.7 L6 19 2.0 

Copyright, TOO. De Loval Steam Turbine Co, Trenton, NJ. 

Fig. 15. — Total-heat-entropy chart for steam. 


Since steam, in flowing through an opening, may be assumed to possess 
no kinetic energy before it reaches the opening. For. (3) will express the 
kinetic energy which it acquires in passing through the opening. For. (1) 
states that: 

(6) Kinetic energy acquired = Heat liberated by adiabatic expansion 
Now, substituting values from Fors. (3) and (5) in For. (6) : 

(7) -^ = 778 W (H: - H2) (ft.-lb. ) 
Or, by transposing and simplifying: 

(8) V = 223.7 V Hi - H2 (ft. per sec.) 

Example. — What theoretical velocity will steam acquire when it is 
expanded through an opening from the dry saturated state at 200 lb. per 
sq. in. abs. to 15 lb. per sq. in. abs. ? Solution. — Use the chart of Fig. 
15 to obtain the values for Hi and H2. In this chart, pressures above 
atmospheric are expressed as gage pressures. Now, 200 lb. per sq. in. 
abs. = 200 — 14.7 = 185.3 lb. per sq. in. gage. Also, 15 lb. per sq. in. 
abs. = atmospheric pressure, closely. Then, from Fig. 15 on the "Dry 
and Saturated Steam" line for 185.3 lb. per sq. in. gage, Hi =1198.1 
B.t.u. per lb. Also from Fig. 15, by adiabatic expansion — following 
vertically downward on Fig. 15 to 15 lb. per sq. in., H2 = 101 B.t. u. 
per lb. Hence, by For. (2) : y = 223.7VHi - H2 = 223.7 X -n/i88.1 = 
223.7 X 13.72 = 3,072 ft. per sec, which is about 15 miles per minute. 

Note. — The Velocity Actually Acquired By The Steam depends 
somewhat on the shape of the opening or nozzle through which the steam 
expands. As the steam flows through the nozzle, the friction of the 
steam on the walls of the nozzle produces heat which is returned to 
the steam and which raises the value of H2 in For. (2), thus reducing 
the amount of heat actually liberated by the steam. In a well designed 
nozzle, that is one in which friction has been minimized by properly 
shaping the nozzle, the actual velocity is usually about 95 per cent of 
the theoretical or computed by For. (2). 

Note. — The Calculation Of Steam Velocities With a Total- 
heat-entropy Diagram, Fig. 15, is much simpler than by using the 
above formula. The entropy (see Practical Heat) is the property of 
steam which does not change during an adiabatic expansion. Therefore, 
the heat liberation during an adiabatic expansion can be traced from a 
total-heat-entropy diagram by following along a vertical (constant 
entropy) line from a selected point representing the initial state of the 
.steam to the line which represents the pressure into which the steam is 
being discharged. The heat liberation is read, as the difference between 
the initial B.t.u. value and the final B.t.u. value, along the vertical scale 
of the diagram. The velocity of the steam (theoretical) can then be 
read from the B.t.u. velocity scale, C, at the top of the main diagram, 
which was computed by using For. (2). 


Example. — By using the chart of Fig. 15, determine the theoretical 
velocity with which steam, at 200 lb. per sq, in. gage and superheated 
100° F., will issue from a nozzle into a region of 29 in. vacuum. Solu- 
tion. — Hi is found, at the intersection of the 100° superheat line and tTie 
200-lb, pressure line, to be 1259 B.t.u. pe lb. Following vertically 
down to the 29-in. vacuum line and then ]C<> the left, H2 is found to be 
863 B.t.u. per lb. Therefore, H: - H2 t 1259 - 863 = 396 B.t.u. per 
lb. From the B.t.u. velocity scale, C, a^^.e top, the theoretical velocity = 
4,460 /^ per sec. (approximately 51 m»'«per minute). 

Note. — The Chart Of Fig. 15 Is Ltrawn For Gage Pressures For 
Use In Districts Where The Average Barometric Pressure is 
about 30 in. of mercury column. Such charts are generally drawn (in 
other books) for absolute pressures; but it is believed that, for most pur- 
poses, one drawn for gage pressures (assuming an average atmospheric 
pressure) will be more useful for the operator. Hence, the one here 
given is so drawn. To use the chart in districts where the barometric 
pressure is much different than that specified above, or for extreme 
accuracy, the reader may calculate the pressure correction required for 
using this chart by the relations: 

(9) Gage pressure value to be used on chart = Actual gage reading — 

[0.49 X (30 — barometer reading)] 

(10) Vacuum gage value to be used on chart = Observed vacuum gage read- 

ing + (30 — barometer reading) 

Or, one may employ a similar chart which is based on absolute pres- 
sures, for example Marks and Davis' Tables and Diagrams of the 
Thermal Properties of Saturated and Superheated Steam (Long- 
mans, Green, and Co.). 

11. The Velocity Of A Fluid Stream May Be Reduced As 
The Stream Passes Over A Moving Surface (Fig. 16). It is 

necessary to know something 
T£,^^£:tZtL'. verify Of about this reduction of veloc- 

Nozz/e (Stationary) '-^B iock^ y/ ^.. ity in Order that One may 

i understand why turbine parts 

,.,: n^^. . ^^^-s™u move at such great speeds 

Velocify Of Water .' /J^ mamt o r- 

leay/ngBiock^s^---' ^^ and why staging (Sec. 17) is 

Ve = Vj -2Vb (Approximately) employed. 

Fig. 16. — Illustrating the loss of velocity 

by a stream of fluid as it does work. EXPLANATION. — In Fig. 6 it waS 

shown how a fluid stream may pro- 
duce a force but, since in Fig. 6 the force of the stream did not move the 
block and hence did no work, the velocity of the stream was undiminished 
(except for a slight loss due to friction) as it passed over the surface of 
the block. But if, now (Fig. 16), the block is free to move, and does 

Se.c. 11] 



move away from the stationary no/zle with a velocity Vb, then obviously 
the stream will not approach the block with its full veloctiy Vj. The 
stream will only strike the moving block with a velocity equal to Vj — Vb,' 
just as. when an automobile going at 45 mi. per hr. overtakes a train 
traveling at 30 mi. per hr.; see also Fig. 17. Thus, if the velocity of the 

Man On Ground Throws Object 
In This Direction With Velocity 
, Of4IHi. PerHr=60Ft Perjec. 

N/ 6round ^ j BBtttltibt^fcjS 

Car Is Moving ZOMi.Per 


I- One Second Later, Object Overtakes Car 

Fig. 17. — Showing why one moving object strikes another only with the difference 
of their velocities. To the man on the platform the object seems to come with a velocity 
of only: 60 — 44 = 16 ft. per sec. or 11 mi. per hr. 

stream is 3,000 ft. per sec. (Fig. 18) and that of the moving block is 
1,400 ft. per sec, the stream overtakes or approaches the block with a 
velocity of 3,000 — 1,400 = 1,600 ft. per sec. 

Now, in passing over the surface of the moving block, the stream 
passes just as fast over a point where it leaves the block as it passes over 

■vj= Jet Velocity 

mo Feet Per Second, s^=Velocity Of B/oc/<^ 
Nozzle--,.. Jl_ J400 n .Per. Second 



yelocify Of 

Leaving / 
Sfeam-^y'bis/ Direction 
lOOrt.PerOf. — -- 



Fig. 18. — Illustrating the velocities of a fluid stream as it strikes a moving surface 
turbine blade) and then leaves in an opposite direction. 

a point where it strikes the block (except for a slight frictional loss). 
Therefore, the stream leaves the block with the same velocity as that 
with which it struck the block, or with a velocity of Vj — Vb to the left. 
That is, in Fig. 18, the stream leaves the block with a velocity of 1,600 
ft. per sec. (with relation to the moving block). 


But, since the block is moving away from the nozzle to the right with 
a velocity of Vb, the actual velocity of the leaving steam jet toward the 
stationary nozzle is made up of — is the difference between — the velocity 
with which the stream passes from the block and the velocity of the block 
itself just as when one throws an object with a velocity of 41 mi. per hr. 
from the rear end of a train traveling at 30 mi. per hr., the object appears, 
to an observer on the ground, to move only at the rate of 11 mi. per 
hr. see Fig. 19. Therefore the stream (Fig. 16) actually leaves the 
block with a velocity toward the stationary nozzle of Ve = (vj — Vh) — 
Vb = Vj — 2vh. Thus, also in Fig. 18, the stream from the moving block 
approaches the stationary nozzle only with a velocity of 1,600 — 1,400 = 
200 ft. per sec. 

Man On Platform Throws Object In 
This Direction With Velocity Of 41 
• Mi. Per Hr. Or 60 Ft Per Sec. 


■ Car Is Moving 30 Mi. 
Per Hr. Or 44- Ft Per Sec. 

I -/^--^j/^^l- Man On Platform Throws Object 

_-Man Catching Object , 

•■^t<---/6^->H< 44 

^.'--""■■■V- Object Goes le-Ft In 
m ! 15ec. = IIMi PerHr. 


^'H-One Second Later, Man On Oround Catches Object' 

Fig. 19. — Showing why, when an object is discharged from a moving body, the actual 
velocity of the object is the difference between the velocity of discharge and the velocity 
of the moving body. To the man on the ground the object seems to come with a velocity 
of only 16 ft. per sec. or 11 miles per hour. 

12. Kinetic Energy Is Converted Into Work As The Velocity 
Of A Jet Is Reduced in passing over a moving surface. The 
truth of this is evident by inspection of For. (3). Applying 
For. (3) to the example of Fig. 16 it follows that: 


Kinetic energy of jet = 



Wherein: W = weight of fluid which passed through nozzle, 
in pounds, vj = the velocity of the jet, in feet per second. 
Also, from Fig. 16, 

(12) Kinetic energy of streain leaving block = -^^ (ft.-lb.) 
Now since, as explained, Ve = Vj — 2vb\ 

(13) Kinetic energy of stream leaving block = — a^ a 




(14) Kinetic energy converted into work = {Kinetic energy of jet) 

— {kinetic energy of stream leaving block) 

Or using the symbols instead of words: 

(15) Kinetic energy converted into work = 

13. A Perfect Steam Turbine Would Convert All The 
Liberated Heat Into Mechanical Work. — Thus, all of the heat 
energy would first be converted into kinetic energy and then, 
in turn, into mechanical work. Obviously, then, for a perfect 
steam turbine: 

(16) Mechanical Work = Heat liberated 

Substituting, then, the expression for the heat liberated 
from For. (5): 

(17) Mechanical Work = TF = 778W(Hi - H2) (ft.-lb.) 
Wherein: W = the work done, in ft. lb. W = the weight of 
steam used, in the ''perfect" turbine, in pounds. Hi = the 
total heat of the steam admitted to the perfect turbine, in 
British thermal units per pound; this may be found from the 
steam chart of Fig. 15. H2 = the total heat of the steam 
after adiabatic expansion to the pressure at which it is 
exhausted from the perfect turbine, in British thermal units 
per pound; this may also be found from Fig. 15 as explained 
under Sec. 10. 

14. The Horsepower And Water Rate Of A Perfect Steam 

Turbine are dependent on the conditions of the steam admitted 

to the turbine and on the pressure at which the turbine 

exhausts steam; the horsepower also depends on the rate at 

which steam is supplied to the turbine; that is, in a sense, on 

the capacity of the boiler. The formulas for the horsepower 

and water rate of a perfect turbine which are derived below, 

follow : 

,.g. o 778W(Hi-H2) W(Hi - H2) ,, , 

^'^^ ^ = 60X33,000 = 2,545 (horsepower) 

(19) W^ = W^%^ (It), per hp.-hr.) 

rli — ±12 


Wherein: P = the power of the perfect turbine, in horsepower. 
W = the rate of steam supply, in pounds per hour. Hi = the 
total heat of the steam under the conditions at which it is 
supplied to the perfect turbine, in British thermal units per 
pound. H2 = the total heat of the steam after adiabatic 
expansion to the exhaust pressure, in British thermal units 
per pound. Wh = the weight of steam flow required for the 
perfect turbine in pounds per horsepower per hour; this is the 
water rate. 

Derivation. — Using For. (17) to express the work done by W pounds of 
steam, there results: 

(20) W = 778W(Hi - Ho) (ft.-lb.) 

But, since in For. (20), W expresses the weight of stsam used in 1 hr., 
W expresses the work done in 1 hr. Now, since a rate of 1 hp. = 33,000 
ft.-lb. per min., also obviously 1 hp. = 60 X 33,000 ffc.-lb. per hour. 
Therefore, to find the number of horsepower in W of For. (20), it is only 
necessary to divide For. (20) by 60 X 33,000; thus: 

,01 N -D 728W(Hi - H2) W(Hi - H2) ,, , 

('') P = 60X33,000 = 2,545 (horsepower) 

which is the same as For. 18. Now, by transposing For. (21) : 

(22) W=^^-^ (Ib.perhr.) 

Ill — XI2 

Dividing For. (22) by P : 

(23) ^ = (h!^- H Jp (Ib.perhp.-hr.) 
But, since W/P = Wh, For. (23) simplifies to: 

(24) Wh = ^ ' „ (lb. per hp.-hr.) 

xli — XI2 

which is the same as For. (19). 

Example. — A theoretically perfect steam turbine is supplied with dry 
saturated steam at 175 lb. per sq. in. gage pressure and exhausts into a 
condenser where the vacuum is 28.5 in. of mercury column. The turbine 
is supplied with steam at the rate of 1,000 lb. per hour. What are its 
horsepower and water rate? Solution. — From the chart of Fig. 15: 
Hi = 1,197 B.t.u. per lb. H2 = 851 B.t.u. per lb. By For. (18): P = 
W(Hi - H2)/2,545 = 1,000 X (1,197 - 851) -^ 2,545 = 136.0 Ap. By 
For. (19): Wh = 2,545/(Hi - H2) = 2,545 ^ (1,197 - 851) = 7.36 lb. 
per hp.-hr. 

Note. — The Theoretical Water Rate Of Any Steam Turbine 
May Be Read From A Chart, AB, Fig. 15. The theoretical water rate 
of any turbine is, of course, the water rate of a perfect turbine operating 

Sec. 15] 



under the same steam conditions. In Fig. 15, values of Wh, as computed 
by For. (19), have been shown in scale A opposite the values of Hi — H2 
on scale B from which they were calculated. The use of scales A and B 
of Fig. 15, therefore, replaces the use of For. (19). 

15. The Horsepower And Water Rate Of A Commercial 
Steam Turbine depend in part on the same factors as do those 
of a perfect steam turbine but, in addition, they depend on the 
efficiency of the turbine in its conversion of heat energy into 
mechanical work. Energy is "lost" in all steam turbines; 
that is, some energy is not converted into work. Some of the 
losses are: (1) F fictional losses at the surfaces over which the 

Brake Horsepower 

Fig. 20. — Graph showing approximate values of the efficiency ratio, based on brake 
horsepower, for commercial steam turbines at rated full load. 

steam rubs. (2) Eddy losses, which are introduced whenever 
the current of the steam suffers an abrupt change in direction, 
as when the steam current strikes anything but the desired 
surface. (3) Windage losses; these are occasioned by the 
motion of the turbine parts within a space which is filled with 
steam vapor. (4) Radiation losses; that is, the heat which is 
lost as such from the hot surfaces. (5) Frictional losses in the 
bearings. (6) Exhaust losses, due to velocity which is not 
converted into work. (7) Leakage losses introduced when 
steam flows through the turbine, or part of it, without passing 
through the desired path. 


The formulas for the horsepower and water rate of a com- 
mercial turbine follow directly from those of Sec. 14 by 
introducing the efficiency and are: 

(25) P^ ^^- ^^"545"'^ (horsepower) 

(26) W. = E.(Hr-H.) ^'^- P^' ^^- ^'■'^ 

Wherein: P^ = the brake (or delivered) power of the 
turbine, in horsepower. E,- = the '' efficiency ratio" or 
efficiency of the commercial turbine as compared with that of 
a perfect turbine, expressed decimally; approximate values of 
Er are given in Fig. 20. Wb = the water rate of the com- 
mercial turbine, in pounds per brake horsepower-hour. The 
other symbols having the same meanings as in Sec. 14. • 

Example. — A 200-hp. commercial steam turbine operates on dry satu- 
rated steam at 175 lb. per sq. in. gage and exhausts into a condenser 
where the vacuum is 28.5 in. of mercury column. What, approximately, 
is its water rate at full load and at what rate must it be supplied with 
steam to develop its full load rating? Solution. — As in the example 
under Sec. 14: Hi = 1,197 B.t.u. per lb. H2 = 851 B.t.u. per lb. From 
Fig. 20, for a 200-hp. turbine, Er = 49 per cent = 0.49. Hence, by 
For. (26): Wb = 2,545/[Er X (Hx - H2)] = 2,545 ^ [0.49 X (1,197 - 
851)] = 2,545 ^ 169.5 = 15.01 Ih. per b.hp.-hr. At 200 hp. the turbine 
will require, as is found by combining and simplifying Fors. (25) and (26) : 
W = Pfi X Wb = 200 X 15.01 = 3,002 lb. per hr. 

16. Theoretically, The Work Done By An Impulsive 
Jet (From A Stationary Nozzle) Is a Maximum If The Velocity 
Of The Moving Surface Which It Strikes Is One-half That 
Of The Jet (for the conditions shown in Fig. 16). The proof 
of this is given below. Actually, the velocity of the moving 
surface must always be slightly less than one-half that of the 

Proof. — It is evident by inspection of Fig. 16 and For. (15) that the 
kinetic energy converted into work will be a maximum when Ve^ =0; 
that is when Ve = 0. Hence, since by Sec. 11, We = Vj — 2%, when /;« = 
then Vj — 2vb = 0. Or, then, by transposing: 2vb = Vj or Vb = Vj/2. 
Hence, it is proved that the work done by the jet is a maximum when 
Vh = Vj/2', that is, when the velocity of the moving surface is one-half 
that of the jet. This result seems logical for (Fig. 16) when Vb = Vj/2 


then Ve = and, then from For. (12), the kinetic energij of the stream 
leaving the block = 0; therefore, the total kinetic energy of the jet has 
been converted into work. 

Note. — The Work Doxe By A Reaction Jet (From A Moving 
Nozzle) Is A Maximum When The Velocity Of The Nozzle Is 
Equal To That Of The Jet. It is obvious that, in order that the actual 
kinetic energy of the steam leaving a reaction wheel (Fig. 1) be zero, the 
nozzle must travel away from the steam jet as fast as the steam passes 
through the nozzle. Then, since no kinetic energy remains in the steam, 
it must all have been converted into work. 

17. "Staging" In A Steam Turbine Is The Division Into Parts 
Of The Processes Of Converting Heat Energy Into Mechanical 
Work. As explained in the previous section, the kinetic 
energy of a jet can be iuhy utilized by conversion into work 
only when the velocity of the moving surface (in an impulse 
turbine; see Sec. 30) is approximateh" one-half that of the jet; 
or, in a reaction turbine (Sec. 31) when the velocity of the 
nozzles is approximately equal to that of the jet. Further- 
more, the velocity of a steam jet is very great (see example 
under Sec. 10 wherein the theoretical jet velocity is 51 miles 
per minute). Such high steam velocities would in actual 
steam turbines necessitate extremely high velocities of moving 
surfaces or nozzles (rotating parts); in fact, structural difficul- 
ties prohibit such velocities except in very small turbines. 
These high steam velocities can, however, be either prevented 
or utilized by ''staging." 

18. There Are Three Classes Of "Staging," velocity, 
pressure and reaction. Each is defined below; see also the 
explanations which follow. 

1. Pressure Staging (Fig. 21) is that in w^hich the conversion of the 
available heat energy of the supplied steam into mechanical work is 
divided into the desired number of steps by causing the steam to expand 
through two or more impulsive-jet nozzles successively or in series, from 
each of which the steam is directed against moving surfaces. There 
will be as many ''steps" (pressure stages, Sec. 40) as there are stationary 
nozzles; in Fig. 21 II there are 4 steps. 

KxPLANATioN. — The Effect Of PRESSURE Stagixg maj" be under- 
stood by a study of the hydraulic analogy shown in Fig. 21. Suppose 
that the level of the water in the reservoir, R, is just 156 ft. above the 
nozzle A. Then water wall issue from A at a velocity of approximately 
100 ft. per sec. Hence, the velocity of the blades or buckets against 


which A directs its stream should, for maximum work, be approximately 
^ of 100 or 50 ft. per sec. (Sec. 16). Suppose, however, that the pressure 
which produces the velocity is divided by the arrangement of Fig. 21 //, 
so that each of the nozzles B, C, D and E is supplied from a tank in 
which the water level is only one-fourth as high above the nozzle as in /. 
Then each nozzle will deliver water at a velocity of approximately 50 ft. 

per sec, and the velocity of the 
blades or buckets of each wheel in 
// need only be 25 ft. per sec. 
Theoretically, arrangements / and 
// will develop the same amount of 
work from a given quantity of water. 
Practically, arrangements / and II 
will give very nearly the same 
amount of work. 

In a pressure-staged steam tur- 
bine, the principle described above 
is exactly duplicated by arrange- 
ments (as are shown in Div. 3) which 
divide the liberation of heat by the 
steam into a satisfactory number of 
steps. The kinetic energy produced 
by each liberation of heat is con- 
verted into work before the next 
liberation of heat. 

Ve/odfi/ OfJef- Vj 
Nozzle- . 

rUns+aged I" Pressure Stage 

( 0ne"5^ep" Or SinglcStage) ( rour"5tep5 Or Four Stages ) 

Fig. 21. — A pressure-staged hydraulic 
turbine. (Analogous to a pressure-staged 
impulse steam turbine.) 

Fig. 22. — Illustrating the principle 
of velocity staging. (Two velocity 
"steps" or stages.) 

2. Velocity Staging (Fig. 22) is that in which the conve rsion of the 
available heat energy, of the supplied steam, into mechanical work is 
divided into the desired number of steps by using a single impulse-jet 
nozzle and then, after the jet leaves the first moving surface, redirecting 
it with guide vanes against additional moving surfaces. There will then 
be as many "steps" (velocity stages. Sec. 39) as there are moving sur- 
faces over which the steam must pass; in Fig. 22 there are two moving 
surfaces, hence two steps. 


Explanation, — The Effect Of Velocity Staging is illustrated in 
Fig. 22. If, instead of being used as in the arrangement of Fig. 16, a 
stream be reversed in direction by a stationary block, A (Fig, 22) and 
thus redirected against a second moving surface on the block, B, the jet 
will again have its velocity reduced by twice the velocity of the moving 
surface. Thus, in Fig. 22, the velocity of the stream as it finally leaves 
the moving block, B, is Ve= Vj — 4 Vh. Hence, for maximum work, Ve = 
and Vb = y,/4. Thus, if Vj = 100, Vb = 25. Comparing this with Fig. 
16, where (Sec, 16) for maximum work vo = vj/2, it is obvious that the 
block in Fig. 22 (which represents buckets on an impulse-turbine rotor) 
need travel only half as fast as that in Fig. 16, for if in Fig. 16, Vj = 100 
then, for maximum work, Vb = 50, 

3, Reaction Staging (Fig, 40) is that in which the conversion into 
work of the available heat energy in the supplied steam is divided into 
the desired number of steps by causing the steam to expand through a 
successive series of two or more moving reactive-jet nozzles. There will 
be as many steps as there are reactive-jet nozzles, 


1. Define a heat engine. Is a steam turbine a heat engine? 

2. Give a brief history of the development of the steam turbine and draw sketches 
to illustrate Hero's and Branca's turbines. 

3. What is the first step in the conversion of heat energy in a steam turbine? Give 
an everyday example of the physical change involved in this first step. 

4. Describe the second step in the conversion of heat energy in a steam turbine. In 
this second step does the action of steam differ from that of any other fluid? Why? 

5. Cite several common examples of impulsive forces. Draw a sketch to show how 
an impulsive force may be measured. What primitive steam turbine utilized impulsive 
forces only? 

6. Give several common examples of reactive forces. Draw a sketch to show how a 
reactive force may be measured. What primitive steam turbine utilized reactive forces 

7. What sort of force is produced when a fluid stream strikes an object and then 
leaves it in an opposite direction? Draw a sketch to show how this force may be 
measured. What kinds of turbines are typical examples of the use of such forces? 

8. How is it shown that steam liberates heat when it flows through an opening from a 
region of high pressure to one of lower pressure? What becomes of this heat? 

9. What relation holds between the kinetic energy which steam acquires in flowing 
through an opening and the heat energy which is liberated? 

10. State the formula for the theoretical velocity of a steam jet. Show its derivation. 

11. How is the actual velocity of the jet related to the theoretical? 

12. Explain the use of the total-heat-entropy diagram for calculating steam velocities. 

13. Explain fully, with a sketch, the reduction of velocity of a fluid stream as it passes 
over a moving surface. What is the relation between the velocity reductien and the 
velocity of the moving surface? 

14. Does a fluid stream gain or lose kinetic energy as it passes over a moving surface? 
Explain fully. 

15. In a perfect steam turbine, what is the relation between work done and heat 
liberated? State as a formula. 

16. What factors determine the horsepower and water rate of a perfect steam turbine? 
State and show the derivation of the formulas. 

17. Explain the use of the chart of Fig. 15 for finding the theoretical water rate of a 
steam turbine. 


18. Name and describe six forms of energy loss in a commercial steam turbine. 

19. State the formula for the water rate of a commercial steam turbine. 

20. Explain fullj' the meaning of the efficiency ratio of a steam turbine. What, in 
general, determines the efficiency ratio of a turbine? What values does it have? 

21. What should be the velocity of a moving surface to insure that a fluid stream in 
passing over the surface will do the maximum amount of work on the surface? Explain 
fully and show the kinetic energy changes. 

22. What should be the velocity of the nozzles of a reaction turbine to provide that the 
steam will do as much work as possible? How is this explained? 

23. How great may the velocity of a steam jet be under some circumstances? State 
why such high steam velocities lead to difliculties in turbines. 

24. Explain how high steam velocities may be either prevented or utilized. 

25. What is the underlying principle of pressure staging? Of velocity staging? 



L. How much heat is theoretically liberated from each pound of steam that flows 
through an opening from a region where the pressure is 175 lb. per sq. in. gage and the 
steam is superheated by 20° F. to a space at atmospheric pressure? 

2. How much heat is liberated (theoretically) when dry saturated at 100 lb. per sq. in. 
gage flows through a nozzle into a region where the vacuum is 28 in. of mercury column 
by gage? 

3. In Prob. 1 what theoretical velocity does the steam attain? Compute by formula 
and compare with result obtained from BC, Fig. 15. 

4. In a perfect turbine, how much mechanical work would be derived from each 
pound of steam in Prob. 2? 

^ 6. If a perfect turbine with steam conditions as given in Prob. 2 consumes 2,000 lb. of 

steam per hour, what are its horsepower and its water rate? Compare the result with 

AB, Fig. 15. 
X 6. What might be expected as the water rate of a 2,000-hp. commercial steam turbine 

which operates under the conditions of Prob. 1 and how much steam would it require per 

hour at full load? 

7. At what velocity should a moving surface (similar to Fig. IG) travel to extract the 

maximum amount of energy from the jet of Probs. 1 and 3? 



19. The Terms Which Are Applied To The Various Kinds 
Of Steam Turbines And To Their Principal Parts will be 

defined and-Ttiuslrated in this division. Terms descriptive 
of turbines and their parts are used with different meanings 
by various writers and manufacturers. It is therefore 
important that the reader understand 
the meanings which will be implied 
by the terms as used in this book; 
hence these definitions. Where sev- 
eral terms are popularly used for the 
same thing, all will be given; the one 
which is preferred and which will be 
used in this book will be stated first. 

Note. — The principal parts of the tur- 
bine will first be defined and illustrated in 
Sees. 20 to 28. Then the various classes 
and types as regards their construction and 
the steam conditions for which they are 
designed will be defined and illustrated in 
Sees. 29 to 46. 


Fig. 23. — De Laval divergent 
nozzles. /-Nozzle used in class 
"C" turbine for high- pressure 
condensing service. JZ-Nozzle 
used in class "C" turbine for 
low-pressure condensing or 
high-pressure non-condensing 

20. A Nozzle (Fig. 23) is an open- 
ing through which steam is allowed 
to flow from a region of high pres- 
sure to one of lower pressure so as 

to acquire additional velocity (Sec. 2). The function of a 
nozzle in an impulse turbine (Sec. 30) is to admit the steam 
to the active or moving parts of the turbine. In a reaction 
turbine, the stationary nozzles admit steam to the moving 
parts which are also of nozzle shape and guide the steam 



from them. Nozzles are usually so constructed that the 
steam flow is restricted by a small opening or throat which 
is the smallest section of the nozzle. The steam is dis- 
charged at the mouth of the nozzle. Steam always expands 

in flowing through a turbine 

. B/acfes Fastened Info Diaphragm 
ySfeam Flow J .'Entrance 

'-Mouth ; '-Throat ' --Uiaptiragm 
■ 'Hozzle Formed Between 
Fig. 24. — Convergent nozzles in 

nozzle; its expansion is at- 
tended with least friction if the 
nozzle is larger where the steam 
enters it than at the throat. 
If the throat of the nozzle is 
smaller than the mouth, the 
nozzle is said to be divergent 
(Fig. 23). Nozzles for large 
pressure-drops are always made divergent. Such divergent 
nozzles, are sometimes called Curtis or De Laval nozzles. If 
the mouth of a nozzle is of the 
same cross-sectional area as the 
throat, the nozzle is said to be con- 
vergent (Fig. 24). Such convergent 
nozzles are sometimes called Rateau 
nozzles. The nozzles of a turbine 
are frequently formed by the open- 
ings between the blades as in 
Fig. 24. 

Not e. — Divergent Nozzles Are 
Sometimes Called "Expanding" Noz- 
zles; and, similarly, convergent nozzles, 
non-expanding. Since expansion occurs 
in nozzles of both types, these terms 

are not strictly correct and should be I- Side View 

avoided. Fig. 25.— Moving blades used 

in class "C" De Laval turbine. 

These blades are formed by the 
21. Blades Or Vanes (Fig. 25) drop-forging process and the bulb 

are curved metallic parts, the ^^^^^^ ^'^ accurately machined to 

. in ^* *^ corresponding recesses in 

function of which is to deflect or the wheel rim. 

change the direction of a cur- 
rent or jet of steam. Blades are sometimes called buckets; 
but buckets are, more properly, the deflecting surfaces of a 
bucket-wheel or tangential-flow turbine (Sec. 43). Blades 



may be either moving blades on which the work of the steam 
is done, or fixed or stationary blades (Fig. 26) which reverse the 
direction of the steam jet so that more work may be abstracted 
from it. Stationary blades are sometimes called guide vanes 

Shroud Ring ■■ 
Fig. 26. — Fixed blades of Allis-Chalmers Parsons turbine. 

or guide blades. The openings between the blades fre- 
quently constitute nozzles as in Fig. 24. 

22. The Rotor Or Runner (Fig. 27) of a turbine is the main 
moving part which carries the blades or buckets. It consists 

Fig. 27. — Complete rotor with two discs. 

mainly of a spindle or shaft which is supported by the bearings 
and which carries one or more discs, D, (Fig. 27) drums (Fig. 
45) or wheels W, (Fig. 31) according to the type of turbine. 
The blades or buckets are carried on the discs, drums or 


23. A Casing, Case Or Housing (Fig. 28) of a turbine is a 
covering — usually a horizontally split cast-iron shell — which 
confines the steam and also acts as a frame for the support of 
the rotor bearings. Casings are frequently provided with 


,'4" Pipe Tap For Steam Seal 
When Operating Condensing 

[<-... ./5/:..»| 'fPipe 

Tap, Drain 




]!■ Bottom Vievv 
Of Foot 

''^"P/pe Drain 
From Gland. 
Til is Is Clean Steam 
And Can Be Led To 
Feed Water Tank 
m-End View 

f" Pipe Open To 

Fig. 28. — Outline dimensions of Steam Motors Company turbine casing, frame No. 5, 
Type B with one bearing pedestal. 

relief valves (Fig. 29) to prevent rupture due to excessive 
pressure. The part of the casing immediately surrounding 
the moving blades, together with the fixed blades and nozzles 

.Center Line Of Er^haust 
^■- Exhaust Casing -.^ \^ Center Line Of Wtieek 

'Adjusting fiut 

Drain-Pipe Connection ' 

I- Sect ion 

H- Front View 

Fig. 29. — Relief valve of Type-6 Sturtevant turbine. This is located on the exhaust 
casing directly opposite the exhaust-pipe opening. 

which it carries is sometimes called the stator (Fig. 30). 

24. A Cylinder (Fig. 30 shows a half cylinder) is a cylindrical 
part of a casing in which a number of the stationary blades of 
the turbine are secured. The term cylinder is most frequently 
used in connection with reaction-type turbines (Sec. 31). 


25. A Barrel {B, Fig. 30 shows the stationary nozzles of 
one barrel) is a group of rotor and stator blades which are 
mounted in rings or drum sections of the same diameter, 
which are the same height, and are so arranged as to act suc- 
cessively on the steam current. There may be a number of 
barrels in one turbine cjdinder. The term barrel is most 
frequently used in connection with reaction-type turbines. 

Fig. 30. — Half cylinder — ^or half stator — of a multi-stage reaction turbine. (Parsons 
type, Allis-Chalmers Mfg. Co.) This turbine has 38 stages. 

26. A Gland {G, Fig. 31) is a device for preventing the 
leakage of steam or air between the stationary parts of a 
turbine and the shaft or the drums which form balance 
pistons. See Div. 5 for further definitions and examples. 

27. A Governor (C and B, Fig. 31) sometimes called the 
speed governor is a device for maintaining the speed of a turbine 
practically constant at all loads; see also Div. 6. Governors 
are either direct governors if the centrifugal force of the weights 
which they employ is the only force used in operating the 
governing valve; or indirect or relay governors if some other 
force is used to operate the governing valve. An over speed 
governor, emergency governor or safety stop {E and V, Fig. 31) 
is a device which operates to stop the turbine when its speed 
exceeds a certain pre-determined value for which the over- 
speed governor has been set; but which is inoperative as long 
as this value is not exceeded. (See Div. 6.) 





Spirvrl Mern'n^hone. 
Gears - ■ ' ' 

Low-Speed Shaff.^ 

28. A Turbine Reduction Gear (Fig. 32) is a mechanism for 
delivering power to a driven machine at a rotative speed con- 
siderably less than the speed of the 
turbine shaft. (See Div. 7.) 

29. Table Showing The Five 
Principal Ways, A To E, In Which 
Steam Turbines May Be Classified 
and the various classes into which 
they are thus divided. The terms 

Conriecfeei To 
Turbine Shaft"' 

;' Driving Pinion— 
which describe these classes will be ^^^- 32.— a turbine reduction gear 

defined in subsequent sections. Study the note on page 30. 


Class with 
respect to 





Blading or nozzle ar- 



Fig. 33 
Fig. 35 
Fig. 83 


Service or steam con- 



High-pressure, non-condensing. 
High-pressure, condensing. 
Low-pressure, condensing. 
Bleeder or extraction. 

Fig. 57 
Fig. 69 
Fig. 79 
Fig. 38 
Fig. 39 







Single pressure and velocity stage 
(axial flow). 

Impulse-re-entry (axial flow). 
Impulse tangential (bucket-wheel). 

Fig. 41 

Fig. 53 
Fig. 42 

Velocity stages 


Single-pressure, several velocity 

Fig. 33 


Reactions tages 


Many reaction stages. 

Fig. 30 

Pressure stages 


Several pressure (impulse) stages 
one velocity stage in each (multi- 
cellular) (Rateau). 

Fig. 67 

Pressure and ve- 
locity stages 


Several pressure stages with several 
velocity stages on one or more 
(Composite) (Curtis and Rateau). 

Fig. 73 

Reaction and ve- 
locity stages 


Impulse-and-reaction turbine. 

Fig. 83 


Direction of flow 


Axial flow. 
Tangential flow. 
Radial flow. 

Fig. 41 
Fig. 42 
P'ig. 43 


Division of flow 


Single flow. 
Double flow. 
Single-and-double flow. 

Fig. 44 
Fig. 45 
Fig. 46 


Note. — Every Turbine Belongs To One Of The Classes Of Each 
Classification, A io E inclusive, shown in the preceding table. For 
example, considering classification. A, every turbine is either an impulse, 
a reaction or an impulse-and-reaction turbine. Also in classification, D, 
it is either axial, radial, or tangential flow. The figure numbers given 
do not, necessarily, indicate the only illustrations in this volume of each 
of the various classes (see also Table 70). In fact, some of the classes 
include a large number of kinds and makes — class No. 17 probably 
includes over half of the steam turbines now in use. The names of the 
various manufacturers who make turbines of these various classes and 
the sizes in which they are made are given in Table 70. 

'c--Uftincf Eye 

Shaft EKfension 


Coupling Casing -■■■'i^ 

Shaft Exfentton 
..r; Glands Governor.. ^_ 

Relief Valve 

2^^ Row Of Stationary Blades/ 
1^^ Row Of Stationary Blades ' 

''Hand Valve For No22le Control 
'Expanding Nozzle 

Fig. 33. — Impulse turbine which has three velocity stages in one pressure stage. 
{Moore turbine.) 

30. An Impulse Turbine, also called a velocity turbine or an 
equal-pressure turbine (Figs. 33 and 34), is one which depends 
almost wholly for its operation on the ''impulsive force" of a 
steam jet or jets which impinge upon the buckets of the tur- 
bine rotor. See Sec. 5 for the definition of an ''impulsive 
force." Thus, an impulse turbine is so designed that the 
expansion of the steam which passes through it — and makes it 



do work — occurs almost entirely in its stationary nozzles or 
in its fixed blades ; practically no expansion of the steam occurs 
in its moving blades. For an impulse turbine, the designer 
intends that the steam jet from the stationary nozzles or 
blades shall impinge on the rotor vanes and thus cause the 
rotor to revolve by virtue of the ^*push" thus produced. The 
usual impulse turbine probably operates about 99.5 per cent, 
by ''impulse" and 0.5 per cent, by reaction. 

Note. — The Pressure Of The Steam Entering The Moving Blades 
Of An Impulse Turbine Is Almost Exactly The Same As That Of 


Nozzle x->\ 




I- Circular Section 

I- Longitudinal 

Fig. 34. — Impulse-turbine blading. 

The Steam Leaving Them. — This follows since there is no expansion of 
the steam in the moving blades; see Fig. 40, / and II. In the nozzles 
or fixed blades, the steam velocity increases as the steam pressure falls 
while in the moving blades the velocity of the steam is expended in 
turning the rotor. 

Note. — The Important Characteristics Of Impulse-type Tur- 
bines are: Few stages, expansion occurs only in stationary nozzles, large 
drop in pressure per stage, best efficiency is obtained when blade velocity 
is appi^imately one half the initial velocity of the steam (Sec. 16). 

31. a' Reaction Turbine, also called an unequal-pressure 
turbine (Figs. 35 and 44), is one which depends principally 
on the *' reactive force" of the steam jets as they leave the 


turbine's revolving blades at greater velocities than those at 
which they approached the blades. See Sec. 7 for the defini- 
tion of a ''reactive force." Thus, a reaction turbine is so 
designed that about half of the expansion of the steam which 
passes through it and causes it to do work occurs in the mov- 
ing blades — and about half in the stationary guide vanes. 
For reaction turbines, the designer endeavors to so design the 

guide vanes and moving 
Nozzles blades that the steam will 

flow into the blades without 
striking them. This he does 
by endeavoring to insure that 
the circumferential speed of 
the moving blades will be 
the same as the velocity of 
the steam stream which enters 

Fig. 35.— Reaction-turbine blading. them. But the moviug bladcS 
The space between the adjacent blades, . 

on the rotor and on the stator, form are SO designed that the steam 
slightly-divergent nozzles. Icaves them at a higher veloc- 

ity than that at which it 
entered them; thus the rotation of the rotor is produced 
by reaction. The usual reaction turbine probably operates 
about 90 per cent, by ''reaction" and 10 per cent, by "im- 
pulse." Hero's turbine Fig. 1 was a purely reaction turbine. 

Note. — The Pressure Of The Steam Entering The Moving 
Blades Of A Reaction Turbine Is Higher Than That Of The 
Steam Leaving Them. This follows because expansion occurs in the 
moving blades; see Fig. 40-///. Some of the heat energy of the steam 
is changed to mechanical work (Sec. 2) in passing through the moving 

Note. — The Important Characteristics Of Reaction-type Tur- 
bines are : Many stages, expansion occurs in moving as well as in station- 
ary nozzles, small pressure-drop in each stage, best efficiency is obtained 
when blade velocity is nearly equal to the highest steam velopity (Sec. 16). 

^ 32. \rhe Distinguishing Difference Between Impulse And 
Reaction Turbines is, therefore, that: In the impulse turbine 
there is no appreciable expansion of steam in the moving 
blades; in reaction turbines there is considerable expansion of 
the steam in the moving blades. Furthermore, it follows 



that: In impulse turbines there is practically no difference- 
between the pressure of the steam which enters the moving 
blades and that of the steam which leaves them; in a reaction 
turbine there is a difference between these entering and leaving 


Fig. 36. 

-On impulse turbines, G\ reads the same as G-i. On reaction turbines Gi reads 

■: Fixed Blades. 


' 'Blading Straightened 
Out Into A Plane 

' • • - -Plane Surface Of Section. 

I- Transverse Section 

E- Longitudinal 


Fig. 37. — Showing relation of a "cylindrical section" to the actual blading of a turbine. 

Note. — To Determine With Pressure Gages Whether A Tur- 
bine Is Of The Impulse Or The Reaction Type, take steam-pressure- 
gage readings Gi and G^, as in Fig. 36. If there is no difference between 
the readings, the turbine is of the impulse or equal-pressure type, because 
in this type there is no pressure drop in the moving blades. If Gi is 
greater than G^, the turbine is of the reaction or unequal-pressure type, 
in which type there is a steam-pressure drop in the moving blades. 

Note. — The Distinguishing Difference Between Impulse Blad- 
ing And Reaction Blading is that the cross-sectional shape of impulse 


blades (Fig. 34-7) is such that the exterior curved surfaces of adjacent 
blades in a row, lie almost parallel to one another whereas the curved 
surfaces of reaction blades are such that the opening between adjacent 
blades is smaller on the steam outlet than on the inlet ride, thus forming 
a nozzle. 

Note. — A Cylindrical Section (Fig. 37) also called a circumferential 
or circular section is employed in Figs. 34 and 35, and in many other 
pictures in this book, in illustrating steam flow in turbine blading. A 
cylindrical section is a section which is taken along a cylindrical surface 
through the turbine blading; it shows as a circle, AAA (Fig. 37) in the 
end view or transverse section. This surface AAA is considered to be 
unrolled or unbent as shown in // and then appears, when looking at it 
from the side, as a "cylindrical section." The blades in a cylindrical 
section all show their true cross sections and spacing, whereas any plane 
section through the blades would cut some of them obliquely and show 
the sections distorted. 

33. An Impulse -and -reaction Turbine (Fig. 83) is one 
which has some of its blading designed and arranged as in an 
impulse turbine and some as in a reaction turbine. See 
Sees. 30 and 31. Some of the largest turbines now in use are 
of this type. The impulse blading is used for the first stages 
as will be explained later. 

34. High-pressure Non-condensing And Condensing Tur- 
bines are turbines which are designed to operate on steam at 
100 to 350 lb. per sq. in. pressure and exhaust at atmospheric 
pressure or into a condenser respectively. The chief structural 
difference between the two is the much larger exhaust steam 
spaces of the condensing turbine which are necessary to 
provide for the large volume of steam at the low pressure of 
the condenser. Non-condensing turbines which are designed 
to operate at a back pressure considerably above atmospheric 
are called hack-pressure turbines. 

^35. A Low-pressure Or Exhaust-steam Turbine is one 
which is designed to operate on low-pressure steam — say 
to 10 lb. per sq, in. gage. A low-pressure turbine is always 
a condensing turbine and usually operates on the exhaust 
from a high-pressure turbine or from a reciprocating engine. 
The low-pressure turbine is characterized by the large steam 
spaces near the admission end which are necessary for the 
large volume which the steam occupies at the low pressure. 
See Sec. 192. 



36. Mixed -pressure Turbines (Fig. 38, also called mixed- 
flow4urbines) are turbines to which steam is admitted at two 
or more pressures. They usually operate on a combination of 
live steam from the boiler and additional exhaust steam from 
some other equipment; this exhaust steam being admitted to 
an intermediate steam belt before the low-pressure stages. 
Thus, steam from both sources flows through the low-pressure 
stages. (See Div. 9.) 

Steam Inlet-' 

Steam Divides Here-, 

Steam Inlet. 

^-- -Exhaust To Condenser 

Fig. 38. — Diagram of a mixed-pres- 
sure turbine. 


Outlet To 

Moving Blades 
■ 'Low-Pressure 
Steam To 
Heating Equipment 

r IG. 39. — Diagram of a bleeder turbine. 

) l^ 

37. A Bleeder Turbine Or Extraction Turbine (Fig. 39) is 

one from which steam is extracted at an intermediate stage and 
led away to be used for some other purpose, usually for heating. 
The usual arrangement is to extract enough steam at about 
atmospheric pressure for feed-water or building heating and to 
allow the rest to flow through the low-pressure stages of the 
bleeder turbine and thence to the condenser. Obviously, 
more steam passes through the high-pressure stages of a bleeder 
tu;d|ine than through its low-pressure stages. (See Div. 9.) 
^^8oA Stage, as defined in general terms, is: A period con- 
stit^tiiig a development or one of several well defined succes- 
sive periods in a development. A steam-turbine stage may be 
defined as a section which comprises, or one of a number of 
well defined sections which comprise, the steam path through 
a turbine. This general definition, however, is indeterminate 
because it does not fix the limits of the section which comprises 


the stage. Different kinds of stages are spoken of in connec- 
tion with turbines but their meanings are not definitely 
estabUshed nor recognized among manufacturers and writers. 
The following definitions are proposed because they are 





1 1 II 



1' II II, II 11' 
Vjr'P''^^^^''^ Reduced In 
X T<"n.'^^f< Of r;^t>^ 


■Pr^sure Reduced h 



1 III 


ThreeSets Of Fixed. 


In One Set Of Fixed 


"^^ In Three Sefs\Of 



Nozzles 1 


u •■■ 














.— 1 










1 ' 



' ' 



' 1 

II , 

veloci t y (jfaph 

1 .1 

1 1 1 

1 1 



1 1 


1 1 




Velocity G'ra 



Velocity GVaph 

IT 1 Ml 

.-Velocity Drops 






















-, Flown ,' M , 

I ];Reactio'n Stages' ' 

I ;i i; ■ ■' " ' 

Casing Pressure Constant' 

'Same Pressure On Both Side's 
Of Each Disc 

1-lmpulse Turbine Consisting I-Impulsc Turbine 

Of Three Velocity Stages 
In One Pressure Stage 

Pressure Falls In Each 

Row Of Blades 

I- Reaction Turbine 
With Three 
Reaction Stages- 

Consisting Of Three 
Velocity Stages 
Grouped Two In One 
Pressure Stage And 
One In A Second 
Pressure Stage 

Fig. 40. — Illustrating different kinds of stages. 

consistent with the most general use of the terms and are 
sufficiently distinct not to be confusing. Different terms 
are used for impulse and reaction turbines because the 
prqcesses are different in the two. 

39^ A Velocity Stage (Fig. 40-/) is that portion of the steam 
path in a turbine, wherein work is done by the impulsive force 


of the steam — see Sec. 5 for definition of 'impulsive force" — 
which consists of one row of stationary nozzles (or one set of 
stationary guide vanes) and the moving blades of the one 
runner which immediately follows the row of nozzles or vanes 
and on which the steam from the nozzles impinges. A 
velocity stage may begin with one row of either nozzles or 
guide vanes and always includes only one set of moving blades. 

40. A Pressure Stage (Fig. 40-//) is that portion of the 
steam path, in a turbine, wherein work is done by the impul- 
sive force of the steam, which comprises one or more velocity 
stages through which the steam passes consecutively, its 
first velocity stage having nozzles and the other velocity 
stages being all which follow up to the next set of nozzles. A 
pressure stage always begins with a set of nozzles but may 
contMn in addition many rows of impulse stationary guide 
^s and corresponding rows of moving blades. 
■ir A Reaction Stage (Fig. 40-///) is that portion of the 
s^~m path, in a turbine, wherein work is done by the reactive 
force of the steam. Sec. 7, which is composed of a set or row of 
stationary nozzles and that row of moving blades upon which 
these nozzles direct the steam. The steam, in passing through 
a reaction stage, suffers a reduction of pressure in both the 
stationary and the moving blades. Reaction stages are 
frequently called pressure stages but it is believed to be better 
to reserve the latter name for the use given in Sec. 40. A half- 
cylinder of a reaction turbine with 38 reaction stages is shown 
in fig: 30. 

42. Various Terms Which Are Used To Designate The 
Staging Of Impulse Turbines and their significance are as 
follows : 

Single-stage Turbines (Fig. 31) are those impulse turbines which 
are composed of but one pressure stage which contains but one velocity 

Velocity-staged Turbines (sometimes called velocity-stage tur- 
bines), Fig. 33, are those impulse turbines which are composed of but 
one pressure stage which contains two or more velocity stages. 

Pressure-staged Turbines (sometimes called pressure-stage tur- 
bines) are those impulse turbines which are composed of two or more 
pressure stages each of which contains but one velocity stage. 


Velocity-and-pressure-staged Turbines are those impulse tur- 
bines which are composed of two or more pressure stages each of which 
contains two or more velocity stages. 

Composite-staged Turbines are those the stagings of which are 
formed as a combination of some of the above stagings, so arranged that 
the steam passes through them in succession or in series: see Sec. 63. 

For a further discussion of staging see Div. 3. 

^7^^ An Axial-fiow Turbine (Fig. 41) is one in which the 
steam flows in a direction approximately parallel to the rotor 

'<i Pipe 

Fig. 41. — Elementary diagram of an axial-flow, single-stage turbine. 

axis. Nearly all large turbines and many small ones are of 
this type. A tangential-flow turbine, also called a hucket-wheel 





Fig. 42. — Elementary diagram of a tangential-flow turbine. 

turbine (Figs. 31 and 42) is one in which the flow of steam is 
approximately tangent to the rim of the wheel. Many small 
turbines are of this type. A radial-flow turbine (Fig. 43) is 
one in which the flow of steam is radially inward toward or 



outward from the shaft. Radial-flow turbines have never 
been regularly manufactured in America but have been 

- Steam Admission - 
•;. Blade Rings-. 

I- Longitudinal Section 

1-Transvers6 Section 

Fig. 43. — Diagram showing action of steam in Ljungstrom radial-flow reaction tur- 
bine. Shafts A and B are forced to rotate in opposite directions; each drives its own 

built in Europe by a Swedish engineer; one is being built in 
the United States. 


; Live-Steam Inlet 
\ ,■ Fixed Blades 

.'Fixed Blades 

Moving Blades^ Live -Steam 

^'^—E/.haust-Steam Outlet 

Fig. 44. — Elementary single-flow reaction 

Fig. 45. — Elementary double-flow re- 
action turbine. 

A Single-flow Turbine (Fig. 44) is one in which nearly 
the steam which drives the turbine flows together 


through the blades in the same general direction parallel to 
the rotor axis. 

45. A Double-flow Turbine (Fig. 45) is one in which the 
main steam current is divided and the parts flow parallel to 
the rotor axis in opposite directions. 

46. A Single -and -double -flow Turbine or semi-douhle-flow 
turbine (Fig. 46) is one in which the steam flows through part 

. , u-LiveSteam 

Steam Divides Mere -^ gYl Inlet 

Fig. 46. — Elementary single-and-double flow turbine. 

of the blades as a single current, then divides into two currents 
which flow in opposite directions. 


1. What are two general classes of nozzles? Name three parts of a nozzle. 

2. In what two ways are blades used in steam turbines? What is the function of the 
guide vanes in an impulse turbine? 

3. What is a rotor? What are its principal parts? 

4. What are some of the functions of a casing? How is rupture of casings by pressure 

5. Define the terms, barrel and cylinder as used in steam turbine nomenclature. 

6. What is the purpose of a gland? 

7. What are speed governors? Relay governors? What other kinds of steam- 
turbine governors are there? 

8. Why is a turbine reduction gear used? 

9. Name four ways in which steam turbines may be classified. Give at least three 
classes under each classification. 

10. What is an impulse turbine? Is it actuated entirely by impulse? 

11. What is a reaction turbine? What other force does it employ? 

12. What is the chief difference between the action of the steam in impulse blading 
and in reaction blading? 

13. What is the direction of flow in a bucket-wheel turbine? What is meant by 
axial flow? 


14. What are pressure stages? Velocity stages? What are the differences between 
the two? 

16. What is the chief difference in design between a condensing and a non-condensing 
high-pressure turbine? 

16. What are the usual sources of steam for a low-pressure turbine? 

17. What is the approximate pressure range for the steam supply for a high-pressure 
turbine? For a low-pressure turbine? 

18. What is the difference between a bleeder turbine and a mixed-pressure turbine? 
To what other equipment may the steam outlets of each be connected. 

19. What is a double-flow turbine? A semi-double-flow turbine? Explain with 




47. Table Showing Classification Of Steam Turbines By 
General Construction. — This classification will be followed 
in the development of this division. Note that this classifi- 
cation combines in one arrangement all of the five classifications, 
A to E, Table 29. 

Blading Or 
Nozzle Ar- 


Staging Or 



(C. Table V6) 


Direction Or 
DIvison Of Flow 

(D And E, Table Z^) 

Usual Steam Conditions 
(B, Table 29) 

Single Stage 

De Laval 

Single Entry 

Ax'ial Flow 

High-Pressure, Condensing 
And Non-Condensing 


Axial Flow 

High-Pressure, Condensing 
Or Non-Condensing-,Or Low- 
Pressure, Condensing 




Single Entry 

Axial Flow 



High-Pressure, Condensing Or 



Non-Condensing; And Low- 
Pressure Condensing 


Curtis And 



Single-And-Double Flow 

High-Or Low-Pres'^ure, 





H.R Element, Single 
Flow- LP Element, 
Double Flow, Or 

High- Pressure Condensing 



Curtis And 

Single Flow 
Double Flow 
Single-And Double Flow 

Hiqh-Or LowPressure, 



Curtis And 


H.t^ Element, Single 
Flow; L.R Element, 
Double Flow. Or 

High-Pressure Condensing 









48v The Three Fundamental Types Of Steam Turbines are, 
see Table 47: (1) Impulse, Sec. 30, and Fig. 50. (2) Reaction, 
Sec. 31 and Fig. 78. (3) Impulse-and-reaction, Sec. 33 and 
Fig. 83. The principal features which are embodied in the 
construction of steam turbines of each of these types are 
described hereinafter in this division. 

Note. — Steam Turbines Are Manufactured In Both The Hori- 
zontal And Vertical Types. In a "horizontar' turbine, the shaft is, 
horizontal. In a ''vertical" turbine, the shaft is vertical. However, 
vertical steam turbines, though formerly widely used, are, except in 
small sizes for driving sump pumps and similar services, becoming obso- 
lete. Step-bearing troubles rendered vertical turbines unreliable. 
Therefore, only horizontal turbines will be discussed in this division. 
The general construction, except bearings, of both types is similar. 

49. The Four Principal Types Of Impulse Steam Turbines 

are (Table 47): (1) Single stage. Sec. 42 and Fig. 50. (2) 




rOe Lava I 
Type Or 
jingle t>tage 

Curtis Type 
(One Pressure 

Two Velocity 

Velocity Staged 
Curtis Type 
tTwo Pressure 
Two Velocity 

TSC- Pressure StagecC 
Or Rateau Type 
(Three Pressure 

Fig. 47. — Illustrating De Laval, Curtis and Rateau types of steam turbines. 

Velocity-staged, Sec. 42 and Fig. 63. (3) Pressure-staged, 
Sec. 42 and Fig. 67. (4) Velocity-and-pressure -staged. Sec. 
42 and Fig. 70. As shown in Table 47 and in the following 
sections, certain of these types may be still further subdivided. 
Also, two types of impulse-turbine staging — usually (2) and 
(3) — may be combined in one turbine. A turbine which is 
made up of such a combination of staging is (Sec. 42) called 
composite-staged . 


Note. — Impulse Steam Turbines May Also Be Classified Accord- 
ing To The Name Of The Man Who Obtained The Original Patents 
(Table 47) as: (1) De Laval. (2) Curtis. (3) Rateau. A single-stage, 
axial-flow turbine (Fig. 47-7) is usually considered to be of the De Laval 
type. Those turbines which belong either to the velocity-staged type 
(Fig. 47-77) or to the velocity- and pressure-staged type (Fig. 47-777) 


valve N 


Fan Casing, 


Fig. 48. — Single-stage 

single-entry turbine direct-connected to a blower, 

manufactured by the Power Turbo-Blower Co. 

are generally called the Curtis type. A pressure-staged turbine (Fig. 
47-7 F) is generally considered to be of the Rateau type. From a con- 
sideration of 7 and IV, it will be noted that, in reality, the Rateau type 
merely consists of a number of turbines of the De Laval type arranged 
in series. It should not be inferred that all of the turbines which are 
manufactured by the De Laval Steam Turbine Co. are of the De Laval 
type. In fact, most of the large-capacity turbines which are manufac- 
tured by this company (Table 70) closely resemble the velocity-staged 
(Curtis) type or the Rateau type. 

Sec. 50] 



■Noiile Valves 

50. The Two Principal Types Of Single-stage Impulse 
Steam Turbines (Table 47) are: (1) The single-entry type, 
Fig. 48, wherein the steam jet strikes the moving blades only 
once. (2) The re-entry type, Figs. 54, 57, and 58, wherein the 
direction of flow of the steam jet is reversed and it is made to 
strike the same set of moving blades or buckets two or more 
times. Different manufacturers' single-stage turbines of 
each of these types are briefly described in the following 

51. The Single-stage Smgle-entry Impulse Steam Turbine 
(Figs. 48 and 49) is the simplest type of turbine. Because of 
their inherently high speeds, mechanical difficulties render 
impracticable the manufacture 
of single-stage single-entry tur- 
bines in capacities greater than 
about 600 hp. If a single-stage 
single-entry turbine is run at 
the proper speed, it is the most 
efficient of any turbine within 
its capacity limits — up to about 
600 hp. However, this proper 
speed is so high, that for most 
services, reduction gears (Div. 

7) will be required. Consequently, it is frequently desir- 
able to run a single-stage single-entry turbine at a speed 
which is much lower than the speed at which it would have 
the maximum efficiency. This is because that, by running 
the turbine at a lower speed, the reduction gear may, for 
these small capacities, sometimes be economically eliminated. 
Turbines of this type are generally designed to operate at 
steam pressures from about 100 to 250 lb. per sq. in., with 
exhaust pressures ranging from about a vacuum of 28 in. of 
mercury up to 35 lb. per sq. in. gage. Their usual operat- 
ing speed is some speed between about 2,000 and 5,000 r.p.m. 
However, some small single-stage single-entry turbines have 
been designed to operate at about 30,000 r.p.m. 

52. The Usual Construction Of Single-stage Impulse 
Turbines Of The Single-entry Type (Table 47) is indicated in 
Figs. 48, 50 and 51 which show turbines manufactured, 

''Valve Chest Exhaust' 

'Steam Inlet 
I-End Elevation I-Sidc Elevation 

Fig. 49. — Coppus steam turbine, type 



JL<----5feofm Inlet 

Shroud T'-. 

Ring ^ '■ Exhaust Pipe 

Fan Castfiof-'' 

Fig. 50. — Single-stage, single-entry Coppus impulse turbo-blower, Type C. (The 
exhaust E, may be so arranged that all, none, or only a part of the exhaust steam will 
be delivered into the blower.) 


%oyernor ^'•Vccuum Breaker 
Fig. 51. — Single-entry, single-stage steam turbine. (De Laval Steam Turbine Co.) 


respectively by the Poioer Turbo-Blower Co., the Coppus 
Engineering & Equipment Co., and the De Laval Steam Turbine 
Co; the method of converting the heat energy of the steam into 
mechanical energy is the same in all of these three turbines. 
The path of the steam through the turbines is, as indicated by 
the arrows, from the inlet, I, through the nozzles, A^, through the 
moving blades, B, and out through the exhaust, E. As the 
steam passes through the diverging nozzles, N, it expands 
(Div. 1). This expansion results in a considerable drop in 
pressure and an increase in velocity of the steam. The pres- 
sure drop is practically equal to the difference between the 
steam pressure at the inlet and that at the exhaust. Practically 
all of the velocity energy which the steam thus acquires is 
converted into mechanical work as the steam jet impinges 
on the moving blades. The steam passes through the moving 
blades only once. 

63. Single-stage Impulse Steam Turbines Of The Re-entry 
Type (Table 47 and Figs. 54 and 57) are but slightly more 
complex in construction than those (Sec. 51) of the single- 
entry type. Because of the fact that the steam strikes their 
moving blades two or more times (see Sec. 18), turbines of the 
re-entry type can be operated with but a slightly lower effi- 
ciency at a much lower speed than can those of the single-entry 
type. Turbines of the re-entry type are, in general, used for 
larger capacities for about the same classes of service as are 
those of the single-entry type. Single-stage turbines of the 
re-entry type are made in capacities of from about 1 to 1,000 
hp. They are designed to operate at steam pressures from 
about 75 to 250 lb. per sq. in., and at exhaust pressures ranging 
from a high vacuum up to about 35 lb. per sq. in. gage. The 
usual operating speed of turbines of this type is some speed 
between about 3,000 and 5,000 r.p.m. 

54. There Are Two Types Of Single-stage Re-entry 
Impulse Turbines (Table 47): (1) Axial flow, Sec. 43 and Fig. 
54. (2) Tangential flow, Sec. 43 and Fig. 57. The principle of 
energy conversion in each type is essentially the same as that 
of the single-entry turbine (Sec. 51). However, in the re-entry 
types only a part of the velocity energy of the steam is given 
up to the rotating wheel the first time it strikes the moving 


blades. After the steam has impinged once on the moving 
blades or buckets it passes through a reversing chamber, which 


Main- \ / 

Governor Valve-' [ '-Nozz/e. p/pef/cfnge 

Emergence/ Valve. 

Fia. 52. — Midwest Engine Company single-stage, axial-flow, re-entry turbine, longi- 
tudinal section. See Fig. 54 for detail of reversing nozzle of thia turbine. 

^Mfrhospheric Relief 




^- -Inlet Yalre 

-Steam Inlet 

Fig. 53. — Cross section of a 10-kw., single-stage, axial-flow re-entry turbine. (lFe«<- 
inghouse Electric & Mfg. Co.) 

changes the direction of flow, whereby the steam is made to 
strike the moving blades a second time. In the axial-flow 

Sec. 55] 



type (Fig. 54) the direction of steam flow is usually changed 
only once; consequently, in this type, the steam-jet strikes 
the moving blades only twice. In the tangential-flow type, 

_^- -Reversing Chamber- - . 

Fig. 54. — Cylindrical section showing arrangement of nozzles and reversing chamber 
of an impulse, single-stage, axial-flow re-entry turbine. (Alidwest Engine Co.) 

Fig. 55.- 

-Showing nozzle and reversing chamber of an axial-flow, single-stage, re-entry 
steam turbine. (Westinghouse Electric & Mfg. Co.) 

the steam jet generally undergoes two or more reversals (Fig. 
58), thus striking the moving blades three or more times. 
Each type is briefly described in the following sections. 

55. The Usual Constructional Arrangement Of Impulse 
Single-stage Re-entry Turbines Of The Axial -flow Type 


(Table 47) is indicated in Figs. 52 and 53. The device which is 
employed to reverse the direction of the steam flow is called the 
reversing chamber. The path of the steam through the nozzles, 
the moving blades, the reversing chamber and back through 


Fig. 56. — Cylindrical section showing arrangement of nozzles which is sometimes used 
on axial-flow single-stage re-entry turbines. (Westinghouse Electric & Mfg. Co.) 

the moving blades the second time is indicated by tne arrows 
in Figs. 54 and 55. Turbines of this type as manufactured by 
the Midwest Engine Co. (Fig. 52) are usually provided with 
three or more nozzles, two or more of which may be controlled 

Nozz/e Valve 


■Terry Type-Z2 turbine, longitudinal section. (Axial adjustment of the wheel 
is made by the wheel nuts.) 

by a hand-operated valve as shown in Fig. 54. The Westing- 
house turbines of this type usually have only one nozzle 
(Fig. 55). However, for certain services, some of the 
Westinghouse turbines of this type are provided with two 
nozzles as shown in Fig. 56. One of the nozzles, H, may be 

Sec. 55] 



.Nozzle Removed 

.-Reversing Buchefs 
Removeof To Show 
Steam Path 

Fig. 58. — Showing path of steam jet in a tangential-flow, single-stage, re-entry tur- 
bine; part of the nozzle and reversing bucket is broken away to better show the steam 
path. See Fig. 59 for the nozzle of a similar turbine. {Terry Steam Turbine Co.) 

y^- Casing 

Reversing Buckets 

'-•Toe Nozzle-' 



Steam Inlet'' 

Fig. 59. — Nozzle and three reversing 
buckets of Sturtevant turbine, made from 
one solid bronze casting. 

■Buc/cet Wheel (Rotor) 

• Fig. 60. — Nozzle valve of Type-6 
Sturtevant turbine. To inspect for 
proper longitudinal alignment of rotor 
and nozzle, remove plug P. The align- 
ment is correct when the edge of 
rotor, R, is flush with the edge of 
nozzle N. 


controlled by a hand valve. The other nozzle, T, is controlled 
by the constant-speed governor. 

56. The Usual Construction Of Impulse Single-stage Re- 
entry Turbines Of The Tangential -flow Type (Table 47)— as 
manufactured by the B. F. Sturtevant Co. and by the Terry 
Steam Turbine Co. — is shown, respectively, in Figs. 31 and 57. 
The principle of operation, as explained below, is the same 
in both turbines. About the only difference between the two 
turbines is in the details of construction. 


Inspection Hole., 

Stuffing Box.-. 

U-— Steam Case 
Nozzle Valve 
^i .■ Bearing Cap 


T-. Collar 
Tripper Mechanism 

Fig. 61. 

'Emergency Valve(Inside) 
-External view of Type-6 Sturtevant turbine. 

Explanation. — The Flow Of The Steam Jet In A Single-stage 
Impulse Turbine Of The Re-entry, Tangential-flow Type is shown 
in Fig. 58. A nozzle and a reversing chamber which contains three 
reversing buckets is shown in Fig. 59. The steam, after passing through 
the expanding nozzle (iV, Fig. 58) strikes the side of one of the semi- 
circular-shaped wheel or rotor buckets. This wheel bucket changes the 
direction of the steam-flow through 180 deg. The steam jet then strikes 
the first reversing bucket, B, of the stationary reversing chamber. This 
stationary reversing bucket again changes the direction of the steam 
flow through 180 deg. so that the steam jet strikes another wheel bucket. 
This reversal is repeated until practically all of the velocity energy of 
the steam is converted into mechanical work of turning the wheel, where- 
upon the steam passes out of the buckets into the casing and then through 
the exhaust. A cross-section of a nozzle valve for, and an external view 
of a Type-6, Sturtevant turbine are shown, respectively, in Figs. 60 and 61. 

57. Impulse Turbine Of The Velocity-staged Type (Table 
47 and Figs. 33 and 64) inherently have a lower rotative 

Sec. 57] 



.'•tnifht Pressure 



Pressure ; 



J2-Pr!e55ure Diagram 

speed than do those of the single-stage single-entry type. 
This is because the velocity-staged turbines employ two or 
three sets of moving blades (Fig. 
62) with a set of stationary blades 
or guide vanes between each suc- 
cessive pair of moving blades. 
The steam is, in the nozzles (iV, 
Fig. 63), expanded from the initial 
pressure (Fig. 63-/7) and tem- 
perature down to the exhaust 
pressure (Fig. 63-/7) and tem- 
perature. About one-half of the 
velocity energy (Fig. 63-///) 
which is thus acquired is, in a 
velocity-staged turbine having two 
rows of moving blades, converted 
into mechanical work in the first 
row of moving blades. After the 
steam has passed through this 
first row of moving blades, the 
direction of flow is reversed by 
the stationary blades so that the 
steam jet strikes the second row 
of moving blades (Fig. 63-//). 

Action Wheel -••. 

Sfaiionary Blades^ 
Or Guide Vanes 

DirecHon Of 
Steam Floir -^^ 

m-VelocUij Dlpgram 





Blades - 

H-Longitudinal Section 

Direction Of F 
Rotation- - 





I-Cyllndrical Section 

Buc/<ef Shroud ' ,■'' 

Blade Or Bucket'' 

Fig. 62. — ^Steam nozzle, revers- 
ing bucket and action wheel of 
Terry turbine. "On action wheels 
the side clearance is the important 
factor. Clearances A and B 
should be kept approximately 
equal and neither should be less 
than ^9 in." 

Y-Veloci + y Triangles 

Fig. 63. — Illustrating action of the 
steam in a velocity-staged turbine 
which has two velocity stages. 


Practically all of the remaining velocity energy is converted 
into mechanical work in this second row of moving blades. 
Consequently a velocity-staged turbine which has two rows 
of moving blades will, when operating under the same condi- 
tions, run at about one-half the speed of a single-stage tur- 
bine. See also Sec. 18. 

Note. — Velocity-staged Turbines Are Not, Inherently, Quite 
As Efficient As Are Single-stage Turbines. Nevertheless, for some 
services, it is frequently desirable to use the velocity-staged turbines. 
This is because that by their use, reduction gears may sometimes be 
dispensed with. For driving a given low-speed unit, a single-stage tur- 
bine with a reduction gear might or might not, depending on conditions, 
provide a higher efficiency than would a direct-connected velocity-staged 
turbine. Velocity-staged turbines, however, generally are simpler, oper- 
ate more quietly, and are lower in first cost than single-stage turbines 
with reduction gears. Consequently velocity-staged turbines find their 
widest application in driving relatively low-speed auxiliaries of small 
capacity where low first cost is of greater importance than is the over-all 
operating efficiency. 

(Outsicle Screen 

(Air Inlet )-._ 

>Noz2le Valve Fan Wheel. '' 


58. In The Usual Constructions Of Velocity-staged Impulse 
Turbines (Figs. 33, 64 and 65) there is one set of expanding 

nozzles with either two rows of mov- 
ing blades and one row of stationary 
blades (Figs. 64 and 65), or three 
rows of moving blades and two rows 
of stationary blades (Fig. 33). More 
than three rows of moving blades are 
seldom used in velocity-staged impulse 
turbines. One method which is em- 
ployed in securing the blades in tur- 
bines of this type is shown in Fig. 
66. For mechanical reasons, velocity- 
staged impulse turbines are only made 
in sizes up to about 1,000 hp. If 
made in capacities much larger than 
1,000 hp., the wheel diameters and 
blade lengths become so great that 
the centrifugal stresses thereby introduced are excessive. 
Their chief application is for driving power-plant auxiliaries 

Air-Outlet Flange 
Fig. 64. — Longitudinal sec- 
tion through Carling impulse 
velocity-staged type, turbine- 
driven volume fan. 

Sec. 59] 



such as centrifugal pumps for circulating cooling water or 
boiler feeding, blowers, small generators and the like. As 
manufactured, their speed ratings and steam service condi- 

"da// Bearing 
Steam Supply 


Fig. 65. — Steam Motors Company turbine, top view with cover removed and certain 
parts shown in section. (Steam Motors Company, Springfield, Mass. See Fig. 113 
for gland details and Fig. 135 for governor.) 

tions are about the same as those given in Sec. 51 for single- 
stage impulse turbines. 

Sfeel-Band ,Electr!calty 
C Shroud-... :'.mided 

I-Section I- Partial Side 

Fig. 66. — Method of attaching buckets to wheel in Moore steam turbine. 

59. Impulse Turbines Of The Pressure -staged Type 

(Table 47 and Fig. 67) consist essentially of several single- 
stage turbines which are contained in one casing and which 
are connected in series. In the pressure-staged turbines 


(Fig. 67) each row of moving blades is separated from the next 
row of moving blades by a diaphragm. This diaphragm con- 

.Overload bypass 

.'Hand nozzle Valve 


Fia. 67. — Axial soction showing gcnorid arranKcinent of a prossuro-staKod turbine which 
has 12 pressure stages. (Z)e Laval Steam Turbine Co.) 


■iSiathnary Nozzles 

'2 Gr More Clearance 

tains stationary blades which are (Sec. 20) of nozzle form. 

The steam, as it enters the tur- 
bine through the first set of noz- 
zles (Fig. 68), is expanded. The 
velocity which the steam thus 
acquires is utilized in doing work 
on the first row of moving blades 
just as was explained in Sec. 52 
for the single-stage turbine. After 
the steam leaves this first row of 
moving blades, it passes through 
the nozzle-shaped stationary 
blades in the first diaphragm. In 
passing through these stationary 
l)lades a second expansion of the 
steam, with a consequent velocity 
increase, occurs. This velocity 
energy is converted into mechanical work in the second row 
of moving blades in precisely the same manner as was 

I[- Circumferential Section 
Fio. 68. — Section of nozzles, buckets 
and wheels of Ridyway turbine. 

Sec. 60] 



explained for the first row. The action of the steam 
throughout the succeeding pressure stages is identical to that 
in either of the first two pressure stages described above. 

/ Note. — The Purpose Of Pressure Staging is to provide a method 
whereby the mechanical difficulties which are encountered in attempting 
to make a single-stage turbine of large capacity may be surmounted. 
The velocity of the steam as it issues from a nozzle is a function of the 
pressure drop (Div. 1). That is, if a large pressure drop occurs, a large 
velocity increase will result, and if only a small pressure drop occurs, a 
correspondingly small velocity increase will result. Therefore, by 
dividing the total pressure drop — inlet pressure minus the exhaust pres- 
sure — into a number of small pressure drops, the velocity with which 
the steam strikes any row of moving blades will be much smaller than 
if all of the pressure drop was produced in one set of nozzles. Conse- 
quently, in a pressure-staged turbine, the velocity and the diameter of 
the rotor can be decreased and the capacity of the turbine increased over 
that of the single- or velocity-staged turbine and yet a comparatively 
high efficiency can be maintained. 

Pneumatic Carbon 
'Governor Packing 

Runner Or Rotor 
Maphragm <• —LiftlncfEyd 

Higti-Pressure Nozzle ■' 'Gland Impeller 

(Water-Sealed Gland) 

Fig. G9. — Section through Ridgway high-pressure turbine. 

60. The General Constructional Arrangement Of Impulse 
Turbines Of The Pressure-staged Type is indicated in Figs. 
67 and 69. Although the principle of operation is the same for 
both of these turbines, the constructional details differ. As 
indicated in Fig. 68, the clearance between the moving and 
stationary parts may be comparatively large. In all pressure- 
staged impulse turbines, some means must be employed to 
minimize the leakage of steam through the clearance between 


the diaphragms and the shaft. This is usually accomplished 
by a labyrinth passageway or by carbon glands (Div. 5). 
To take care of overloads, the turbine may be provided with 
extra nozzles which may be either hand- {H, Fig. 67) or 
governor-controlled, or they may be provided with a bypass 
(B, Fig. 67) so that high-pressure steam may be admitted 
directly to one of the intermediate stages. Turbines of this 
type are usually designed for capacities of from about 500 to 
5,000 kw., to operate with either high- or low-pressure steam, 
condensing or non-condensing, at some speed between about 
3,000 and 5,000 r.p.m. 

61. Impulse Turbines Of The Velocity-and -pressure- 
staged Type (Table 47, and Fig. 70) consist, essentially, of a 

Nozzle diaphragm, 3rd stage.^ ,'2ncf Stage guide vanes 

3rd. Stage _ 
■guide vanes 

/■Nozzle diaphragm, 2nd stage 
,1st Stage guide vanes 


d^'^'P SrdSfagi \ , ^ J \ 
Wheef l^rid Stage \ 

Diaphragm packing^' ^''°'" 

Inlet- •' 

^Packing drain 

'Isf. Stage drain 
^Isf.Stage wheel 
2nd Stage wt^eel 

Fig. 70. — Sectional view showing assembly of a velooity-and-pressure-staged impulse 
turbine which has three pressure stages with two velocity stages in each pressure stage. 
(General Electric Co.) 

number of velocity-staged turbines which are contained within 
the same casing, and which are connected in series. The 
total steam-pressure drop between the inlet and exhaust is 
divided into a number of smaller drops as in the pressure- 
staged turbine (Sec. 59). Then the action of the steam in 
each pressure stage is the same as that which was described 
(Sec. 57) for the velocity-staged turbine. The purpose of 

Sec. 62] 



velocity-and-pressure staging is about the same as that of 
pressure staging (Sec. 59). Turbines of this type are also 
sometimes called the Curtis type (Sec. 49). 

62. The General Construction Of Impulse Turbines Of The 
Velocity-and-pressure -staged Type is illustrated in Fig. 70. 
Turbines of this type are made in capacities of from 10 to 400 
hp. and are adapted to operate at pressures from 60 to 250 lb. 
per sq. in., condensing or non-condensing, at some speed 
between about 1,200 and 5,000 r.p.m. The non-condensing 

.•Hoving Blades-^ 

\ Sfafhnary \ 

'> Guide Vanes- ; 



^Stationary Nozzles,^ 

'<---^+age" — -^- ^«+^°" 5tage5--..->l 

Fig. 71. — Cylindrical section through nozzles and blades of a composite-staged steam 
turbine. Five stages are shown. {Moore Steam Turbine Corp.) 

units of this type have two pressure stages. The condensing 
turbines have two, three or four pressure stages, depending 
upon the capacity and upon the operating conditions. Each 
pressure stage has two rows of moving blades and one row of 
stationary guide vanes. Diaphragms separate the pressure 
stages from each other. These diaphragms are provided with 
nozzles, just as are the pressure-staged turbines. Each 
diaphragm is provided with a metal labyrinth packing to 
minimize steam leakage along the shaft. Those turbines of 
this type which have three or four pressure stages may be 
arranged for either mixed pressure or extraction service 
(Div. 4). 


63. Impulse Turbines Of The Composite-staged Type 

(Table 47 and Fig. 71) usually consist of a number of pressure 
stages. The first pressure stage (Fig. 71) usually contains two 
velocity stages. This first stage is followed by the required 
number of pressure stages, each of which contains one ve- 
locity stage (for exception see Fig. 76). The first stage is 
sometimes called a Curtis stage, and those which follow are 
sometimes called Rateau stages. Therefore, a turbine of the 


p.erM Byp.S5 Pfpes .'■■^fJSh^i: 



Fig. 72. — Partial longitudinal section of a high-pressure composite-staged impulse tur- 
bine which has twelve pressure stages. {General Electric Co.) 

composite-staged type is, in reality, a velocity-staged turbine 
(Sec. 57) which has in series with it a pressure-staged tur- 
bine (Sec. 59). The action of the steam through such a 
turbine may be understood from a study of Sees. 57 and 59. 

Note. — The Reasons For The Use Of Composite Staging in impulse 
turbines are that, for the larger capacities — above about 1,000 kw. — 
they are more efficient and less expensive to construct than turbines of 
any of the types which are described in the preceding sections. This is 
because the two velocity stages, which are in such turbines always placed 
in the high-pressure end, will efficiently cover an expansion range equal to 
several pressure stages. Thus, by employing them, the size and conse- 
quently the cost of the turbine may be reduced. Also, by placing the 
two velocity stages in this first pressure stage, the pressure of the steam 
therein may be considerably reduced over that which would be required 
if the velocity staging were replaced by equivalent pressure staging. 

Sec. 64] 



This decreases the windage loss and the leakage of steam, thereby 
increasing the efficiency. 

re/to f Valve 

Fig. 73. — Longitudinal section of an impulse turbine of the composite-staged type 
having one Curtis and five Rateau stages. {I nger soil-Rand Co.) 

^ Lifting Eye 
Relief Valve 
,M9ving Blaoles 

Coil ■ 

''■£xhcxusf riancfe' 

Fig. 74. — Longitudinal section of a composite-staged impulse turbine. {Terry Steam 

Turbine Co.) 

64. Various Methods Of Construction Of Composite-staged 
Impulse Turbines are illustrated in Figs. 72, 73, 74, 75 and 


76. The same principle is employed in all of these turbines. 
However, the different manufacturers follow different 
mechanical designs. Practically all impulse turbines between 

Eye Bolt 


End Cast/ngf - - -j^ 

Cavity — " 


(Diaphragm Cover 

5team-Enc( Casting 

'Impulse Bucket 

''Oil Baffle 
■Steam Supply Valve 

'Live Steam Cavity 
''Nozzle Valve 

Fig. 75. — Sectional view of Moore steam turbine. (Instruction Card No. 1, Moore 
Steam Turbine Corporation, Wellsville, New York.) 

ilYtelve RateauStages. Two Curtis ^^-^ 

'• Stages^ Governor- 

Exiyaust Flange 

Fig. 76. — Longitudinal section through a 1,000 kw. Kerr Curtis-Rateau type turbine — 
two Curtis stages and twelve Rateau stages. 

about 1,000 and 35,000-kw. capacity are of the composite- 
staged type. However, they are also regularly manufactured 
in smaller capacities. Their usual operating speeds are from 
about 1,500 to 5,000 r.p.m. They are made for high, low 

Sec. 65] 



and mixed pressures, condensing and non-condensing, and 
(Div. 9) for extraction service. In general, they are used to 
drive large-capacity generators. 

65. In A Reaction Turbine The Steam Expands In Both The 
Moving And In The Stationary Blades (see Sec. 31 and Fig. 77). 

Fig. 77. — Obsolete Allis-Chalmers reaction-turbine blading. 

The steam is admitted to the first row of nozzle blades {E, Fig. 
78) at full inlet pressure. The steam, in passing through these 
blades, undergoes a slight expansion. A further expansion of 

High Pressure section 

low Pressure Section 

By-pcrss Vafve 
Baiance Piston^- 

Fia. 78. — Longitudinal section of a single-flow reaction turbine. (Allis-Chalmers Mfg. 


the steam occurs in the moving blades; the work of rotation is 
thus produced by reactive forces (Sec. 7). The action of the 
steam in each successive reaction stage of a reaction turbine is 
identical to that in the first reaction stage which is described 
above. To take care of the increasing volume of the steam 


as it expands throughout the successive reaction stages of 
a reaction turbine, the blade lengths and the rotor diameter 
are increased by successive steps (barrels, Sec. 25) or sections 
(H, J, and K, Fig. 78). A reaction turbine is sometimes 
called the Parsons type of turbine. 

66. Reaction Turbines Are Manufactured For Nearly All 
Steam Conditions. However, they are usually designed for 
operation on high- or low-pressure steam and to exhaust into a 
high-vacuum condenser. The reason for the general use of a 
high-vacuum condenser with turbines of this type is that the 
intermediate and low-pressure sections ( Fig. 78) are more 
efficient than is the high-pressure section. The most efficient 
speeds of reaction turbines are usually lower than are those of 
impulse turbines because reaction turbines are generally built 
with very many more stages. Consequently, they are generally 
used to drive large alternators through a direct connection, at 
some speed between about 750 and 3,600 r.p.m. Reaction 
turbines are made in capacities of from about 300 to 30,000 
kw. For capacities above about 30,000 kw., compound turbines 
(Sec. 68) are generally used. Various types of reaction turbines 
are described in the following sections. 

67. The Three Principle Types Of Reaction Turbines 
(Table 47) are: (1) Single-flow, Sec. 44 and Fig. 78. (2) 
Double-flow, Sec. 45 and Fig. 79. (3) Single-and-double-flow, 
Sec. 46 and Fig. 80. Reaction turbines of each of these 
types are described in the notes below. 

Note. — A Single-flow Reaction Turbine is shown in Fig. 78. The 
live steam is admitted through the inlet, C, to the high-pressure section, 
H, of the cylinder at E. After passing through the turbine, the steam 
is exhausted at G. The difference in the steam pressure — which is 
caused by the expansion of the steam in the moving blades, Sec. 65 — on 
the two sides of each row of moving blades produces an end thrust in 
the direction (to the left in Fig. 78) of the steam flow. To equalize this 
end thrust, balance pistons, L, M and A'^, are provided, respectively, for 
each of the three sections //, J and K. These pistons connect with the 
high-pressure ends of their respective sections by the passageways E, O 
and P. The area of the balance pistons, L, M and N, is just sufficient 
so that the steam pressure on them exactly balances the end thrust to 
the left. To operate at overload, the govQrnor-controlled bypass valve, 
V2 (Sec. 154), admits steam directly to the intermediate-pressure sec- 
tion J. 

Sec. 67] 



Note. — In A Double-flow Reaction Turbine (Fig. 79), the steam 
is admitted at the center of the blading at A. There the steam divides 
into two equal parts. One-half of it flows to the left and the other half 
flows to the right. Consequently the end thrust (see preceding note) in 

Exhaust / f Lv5Zitk__JJ_ 
Space " 

Fig. 79. — Low-pressure double-flow reaction turbine. {W estinghouse Electric & Mfg. 


one-half of the turbine is counter-balanced by that in the other half, 
thus obviating the necessity of balance pistons. Also, since the steam 

Relief Va/vQ^ 


.Relief Va/re 



Fig. 80. — Section of a 30,000-kw. single-double-flow steam turbine having reaction 
blading and complete expansion within a single cyhnder. Note that the legend Over- 
load Admission Spces should read Overload Admission Spaces. (Westing house Electric & 
Mfg. Co.) 

is divided into halves, the diameter of the rotor can, in a double-flow 
turbine, be made smaller than in a single-flow turbine of equal rating. 

Note. — A Single-and-double-flow Reaction Turbine (Fig. 80) 
provides a means of utilizing the energy in the large volume low-pressure 


steam without unduly increasing the blade lengths in the low-pressure 
sections of the cylinder. In Fig. 80, the steam enters the turbine at the 
admission space, S, and flows through the turbine, in a left-hand direc- 
tion, to A. At A, the steam divides, one-half flowing through the low- 
pressure section B and the other half, via NMP, through the low-pressure 
section C, to the exhaust. 

5fec/m Passacfe From 
High-To Low- Pressure- 

■ Hlgh-PressuKe 


■Lxhausf To 

iiJ \:5feam 


Fig. 81. — Exterior view of a tandem-compound reaction turbine; the high-pressure 
cylinder, H, is of the single-flow type, the low-pressure cylinder, L, of the single-and- 
double-flow type. See Fig. 245A for a sectional elevation of a tandem-compound reac- 
tion turbine. 

.Alfernafors, lO.OOO-Kv/. Eac^ 


Fig. 82. — Three cylinder cross-compound, 50,000-kw. reaction turbine unit consist- 
ing of one high-pressure and two low-pressure elements. (Westinghouse Electric & Mfg. 

68. A Compound Steam Turbine is one wherein the total 
steam expansion from boiler pressure to condenser pressure 
occurs in two or more separate cylinders. Compound steam 
turbines are (Table 47) made: (1) Tandem-compound, Fig. 81, 
wherein the axes of both cylinders lie along the same straight 
line. A tandem-compound turbine unit is usually direct-con- 

Sec. 69] 



nected to a single generator. (2) Cross-compound, Fig. 82, 
wherein the axes of all cylinders are not in the same line, 
but usually in parallel lines. Each element, or cylinder, of 
a cross-compound turbine unit is usually direct connected to a 
separate generator. The tandem-compound reaction turbine 
which is shown in Fig. 81 has a high-pressure cylinder of the 
single-flow type and a low-pressure cylinder of the single-and- 
double-flow type. 

69. An Impulse -and -reaction Turbine (Fig. 83) is, in addi- 
tion to the reaction blading, R, generally provided with two 

Exhaust flange. 

Fig. 83. — Single-flow impulse-and-reaction turbine of 10,000-kw. capacity. {.Westing- 
house Electric & Mfg. Co.) 

rows of moving blades, V, of the velocity-staged impulse type 
(Sec. 57). The steam flows through this impulse blading 
before it reaches the reaction blading. Thus both the tempera- 
ture and pressure of the steam is decreased before it enters the 
first reaction stage. Since the steam pressure on the first 
reaction stages is thereby decreased, the leakage of steam over 
the ends of the short reaction blades will not be as great as if 
the high-pressure steam were admitted directly to the first 
reaction stage as is done in turbines (Sec. 65) of the purely 
reaction type. Also, since the temperature of the steam is, in 
the impulse-and-reaction turbine, lowered before it reaches the 


reaction blading, the high-pressure section of an impulse-and- 
reaction turbine is not subjected to such high temperatures as 

hmliary Steam Infef 


Fig. 84. — Section of a 20,000-kw., 1,500-r.p.m., double-flow Westinghouse impulse-and- 

reaction turbine. 

is the high-pressure section of a purely reaction turbine. 
For the same reasons as were given for reaction turbines (Sec. 

Valve '■ 

FiQ. 85. — Westinghouse single-and-double flow, impulse-and-reaction 25,000-kw. turbine. 

67), impulse-and-reaction turbines are made single-flow (Fig. 
83), double-flow (Fig. 84), and single-and-double-flow (Fig. 85). 



1. Name the three fundamental types of steam turbines. 

2. What is a horizontal turbine? A vertical turbine? What tends to cause vertical 
turbines to be unreliable? 

3. What are the four principal types of impulse steam turbines? What two of these 
types are frequently combined into one turbine? 

4. Name the classification of impulse turbines according to the name of the man 
obtaining the original patent. Make a sketch to illustrate each. 

5. For what purposes are single-stage impulse steam turbines generally used? What 
is about their usual maximum horsepower rating? Range of pressure ratings? Range 
of speed ratings? 

6.' Name two principal types of single-stage impulse steam turbines and explain 
with a sketch the action of the steam in each type. 

7. Name the two principal types of single-stage re-entry turbines and make a sketch 
to show the path of the steam through each type. 

8. For what classes of service are single-stage impulse re-entry turbines especially 

9. Explain with a sketch the action of the steam in a turbine of the velocity-staged 
type. Does the velocity of the steam with respect to the vanes or blades change in 
passing through them and if so how? 

10. What is the maximum number of rows of moving blades which is generally used in 
velocity-staged turbines? 

11. Why are velocity-staged turbines sometimes used in preference to single-stage 

12. Make a sketch to show the usual arrangement of the nozzles, moving and sta- 
tionary blades in a velocity-staged turbine. 

13. What are the principal applications of velocity-staged turbines? For what 
speeds, horsepowers and steam conditions are they usually designed? 

14. Make a sketch of and explain the action of the steam in a pressure-staged impulse 

16. What is the purpose of pressure staging? Explain how pressure staging accom- 
plishes this purpose. Has pressure staging any advantage over velocity staging and 
if so what is it? 

16. What is a diaphragm? What means are generally employed to minimize steam 
leakage through the clearance between the diaphragm and the shaft? 

17. What two methods are used on pressure-staged turbines to provide for overload? 

18. Give the horsepower range, the usual steam conditions and the speed range for 
which pressure-staged turbines are ordinarily designed. 

19. Make a sketch to illustrate the action of the steam in a velocity-and-pressure- 
staged turbine. What is the purpose of velocity-and-pressure staging? 

20. What are the horsepower range, the usual steam conditions and the speed range 
for which velocity-and-pressure-staged turbines are usually designed? 

21. Make a sketch to explain the action of the steam in a composite-staged turbine. 
What is the reason for using composite staging? Within what horsepower and speed 
ranges are composite-staged turbines usually designed to operate? 

22. Explain the action of the steanl in a reaction turbine. 

23. Give the range of speed and horsepower ratings for which reaction turbines are 
ordinarily designed. For what steam conditions are they especially suitable? 

24. What are three principal types of reaction turbines? 

26. Why are balance pistons generally used in single-flow reaction turbines? Why 
are they not required in double-flow reaction turbines? 

26. Why is the single-and-double-flow construction used in large reaction turbines? 

27. What is a compound turbine? What is a tandem-compound turbine? What is a 
cross-compound turbine? 

28. What are the advantages of an impulse-and-reaction turbine over a reaction 

29. Make a complete table showing the classifications of all steam turbines according 
to general construction. 

70. Table by Manufacturers Showing Steam Conditions, 



Steam conditions 

Class or description 

Alberger Pump & Condenser Co. 

High-pressure, non- 

Curtis, impulse, sin- 
gle flow 

Allis-Chalmers Mfg. Co 



Parsons, reaction, sin- 

Calling Turbo-Blower Co 


High-pressure, non- 

Impulse, single-flow 

Coppus Engineering & Equip- 
ment Co. 


High-pressure, non- 

Impulse, single-flow 


High-pressure, con- 
densing and non- 

Impulse, single-flow 

De Laval Steam Turbine Co 

High-pressure, con- 
densing and non- 

Impulse, single-flow 



Impulse, single-flow 


High-pressure, non- 

Curtis, impulse, sin- 



Curtis, impulse, sin- 


High- and mixed- 
pressure, condens- 

Impulse, single-flow. 

and double-flow 




High-pressure, non- 

Impulse, re-entry 

Moore Steam Turbine Corp 

High-pressure, con- 
densing and non- 

Impulse, single-flow 


Parsons Marine Steam Turbine 

High- and low-pres- 

Parsons, reversing 

* Oil relay governors used on some large machines. 

t Steam relay governors made at Lynn works; oil at Schenectady. 


Classes and Approximate Ratings of Steam Turbines 


Type of 




10-50 hp. 




Centrifugal-pump drive 

300 kw. 
and up 


Throttling and by- 

Water packed 

Turbo-generator, direct 

1-25 hp. 


Direct throttling 

Metal packing 

Blower drive 

2-50 hp. 



Stuffing box 

Blower and pump 

1-600 hp. 


Direct throttling 

Metal packing 

1-600 hp. 


Direct throttling 

Metal packing 

Direct or gear-con- 
nected for pump or 
generator drive 



Direct throttling* 
or oil relay 

Carbon packed 



Direct throttling 

Steam packed 

Mechanical drive 

100 kw. 
and up 


Direct, t steam, oil 

Steam laby- 

Director gear-con- 
nected for generator 



Oil relay 

Steam laby- 


Up to 
4,500 hp. 


Direct throttling* 
and oil relay 

Carbon packed 

Direct and gear-con- 
nected for pump and 
generator drive 


1-800 hp. 


Direct throttling 

Carbon packed 

Direct connected for 
pump, generator, and 
blower drive 

Up to 
4,500 hp. 


Direct throttling 
and oil* relay 

Carbon packed 

Direct and gear-con- 
nected for pump, gen- 
erator and blower 



Marine service 

70. Table by Manufacturers Showing Steam Conditions, Clas- 



Steam conditions 

Class or 

Power Turbo Blower Co 




Impulse, single-flow 

Ridgeway Dynamo & Engine Co. 


High- and mixed- 

Impulse, single-flow 




Impulse, single-flow 

B. F. Sturtevant Co 




Impulse, tangential- 

Terry Steam Turbine Co 




Impulse, tangential- 




Impulse, single-flow 

Westinghouse Electric & Mfg. 




Impulse, re-entry 


82, 83, 


Single-flow, double- 
flow, single-and- 
double flow, single- 
cylinder and com- 

L. J. Wing Mfg. Co 




Oil relay governors used on some large machines. 


ses, and Approximate Ratings of Steam Turbines — Continued 


Type of 




1-20 hp. 



Metal packing 

Blower drive 



Pneumatic oil re- 

Carbon pack- 

Generator drive 

Up to 
300 hp. 


Direct throttling 

Steam laby- 

Direct-connected pump 

5-350 hp. 


Direct throttling 

Stuffing box 

Direct pump and gen- 
erator drive 



Direct throttling 

Carbon packed 

Direct and gear-con- 
nected for pump, 
blower and generator 



Direct throttling 
or oil relay * 

Carbon packed 

2-640 hp. 


Direct throttling 

Water seal 

Mechanical drive 

500 kw. 
and up 

Reaction, im- 

Steam or oil re- 
lay, and inter- 



1-20 hp. 



Stuffing box 

Blower drive 


71. The Various Steps In Installing A Steam Turbine may 
be given in the order in which they should usually be performed 
as follows: (1) Plan if necessar\\ The principal parts of the 
installation which must be planned are the foundation and 
piping. (2) Build the foundation. (3) Receive and unpack 
the turbine. (4) Pluce the turbine on the foundation, level and 
grout. (5) Make initial adjusttyietits. The bearings, coupling, 
governor and thrust bearings should be adjusted suflSciently 
so that the turbine may be turned over slowly without damage. 
(6) Conned up the condenser, oil system, piping, drains, and 
other accessories. (7) Make final adjustment under operating 
conditions. (8) Start up the first time. The governor must 
be adjusted by running the turbine at its rated speed. 

72. In Planning The Installation Of Large Turbines, 
(Fig. 86) pro\'ision should be made for the space and support 
required by all principal piping, bearing in mind that the 
turbine casing must not be subjected to piping strains. The 
location of all auxiharies and accessories should be carefully 
planned so that they can be readily handled b}' the crane 
and so that they are all as nearly alike as possible thus facilitat- 
ing the stocking of spare parts. The method of cooling the 
generator, of supporting the condenser and of connecting the 
turbine exhaust passage should be completely planned. 
The planning of large tm-bine foimdations and supports 
involves pro\4sion for the extra stresses occasioned by the 
the vacuum in turbine casing. (See the author's Machinery 
ForxDATioxs AXD Erectiox.) 

Note. — The Piping For A Small Turbixe (Fig. 87) need not ordi- 
narily be accurately planned. The turbine may be located where desired 
and a pipe Une run to it and exhaust line run from it by an experienced 
steamfitter but it must then be properly supported to reUeve .:ie turbine 


Sec. 73] 



casing of all stresses. Provision is sometimes made in small turbine 
piping for special governing. Pressure-controlled diaphragm valves are 
sometimes used on turbo-blowers for boiler furnaces (Fig. 87) so that the 
speed of the turbine will be proportional to the steam requirements of 

-Steam Supply-'' 

Outlet Air Dud 

Free Exhaust To , . , , ^_ 

AfmospHere,^^^^.^^ K^^J 
/ Exhaust- - ' - 


Exhaust Reiki 


Surfoce Condenser 

Lire , A'lr 

Steam \ '[Elector 

Clrcukiiing Pump: Supply If ate/ Make-ip' Hot-fi'ell Pump- • 'Hydraulk S^pp.y pjrr:p 
Fig. 86. — Turbo-generator installation showing principal auxiliaries and piping. 

the plant — if the boiler pressure falls, the turbine will furnish more air 
and %'ice versa. Similar valves or pump governors are sometimes used 
on turbine-driven boiler-feed pumps to keep a desired water pressure in 
the feed line. 

To Boiler Pressure--, To Steam 
ReyuMIn^ VaAie.^ \^^[\ 
Turbine •■_ Pressi/re. h 



Fig. 87. — Piping connections, for turbine-driven blower, which enable blower to main- 
tain a constant steam pressure in the boiler which it sers^es. 

73. Foundations For Large Or Medium-sized Steam 
Turbines are ordinarily built in hollow form so that the 
condenser and other auxiliaries may be placed directly beneath 


the turbines. The foundation is frequently in the form 
of a raised platform or capslab supported from a subbase or 
footing by a superstructure consisting mainly of vertical 
columns. Foundations for small non-condensing turbines 
are simple blocks of concrete which differ in no essential 
respect from foundations for small motors or other small 
machinery. In no event, however, should a turbine bedplate 
be bolted to a wooden floor without building for it a suitable 
rigid concrete slab (or structural steel frame for small turbines) 
which will protect the bedplate from possible bending. It 
should be remembered that the function of the foundation 
is to maintain the unit in alignment. Permanence of align- 
ment is largely dependent upon the rigidity of the foundation. 

74. In Receiving And Unpacking A Turbine Which Is 
Shipped Assembled (from General Electric Co. ^'Instruction 
Book 82,200") see that the blanks over the piping outlets and 
inlets are intact and that no foreign material has lodged in 
the steam passages. Look over the gages and piping and see 
that all the fittings are in place. Report any shortages as 
soon as possible. When ready to install the machine, wipe 
off all slush with clean waste and, if carbon packing boxes or 
other machined surfaces coming in contact with the steam or 
exposed to view or touch are slushed, they should be cleaned 
thoroughly with gasoline. No bearings, linings, journals, or 
roller or ball thrusts should be cleaned with gasoline but merely 
wiped clean with waste. 

75. In Receiving And Unpacking A Turbine Which Is 
Disassembled, locate all parts called for on the shipping 
memorandum. If the parts are to be assembled immediately, 
wipe off all slush or clean with gasoline as previously noted. 
The wheels and shaft will, in most cases, be shipped resting in 
blocks fitted to the recesses in the heads where the carbon 
boxes belong and this blocking should not be removed until 
the wheel casing is resting in the base and the shaft in the 
linings. See that all the blanks over the openings are intact 
and that no foreign material has found its way into any part 
of the machine. 

76. Turbines Are Placed On Their Foundations And 
Aligned On Wedges (Fig. 88). The wedges are of steel about 

Sec. 77] 



1 in. thick. The primary aUgnment and leveHng is usually 
made with all principal parts of a turbine in place but before 
the piping and auxiliaries are connected. The machine is 
slid slightly or the wedges driven in or out until the desired 
level of the bedplate is obtained. The level is indicated by 
placing an accurate spirit level across the finish bosses of the 
bedplate. These bosses are usually provided at convenient 
points on the bedplate and are scraped to an accurate level at 
the factory. It is not sufficient to try the level at one or two 



- C7.-:.A^-^?;, 

Fig. 88. — Bedplate of turbine supported on wedges and surrounded by wooden dam 

for grouting. 

points. It should be tried all the way around since there is 
often some warping of the bedplate in shipment. The bed- 
plate is then grouted to the foundation by pouring thin grout 
or cement mortar under the plate. A dam is built (usually 
of strips of wood) to confine the grout and force it to fiow under 
the plate and up inside for 2 or 3 in. After about 2 or 3 hr. 
the dam is removed and the excess grout trimmed off. About 
two days later, the wedges may be removed, if desired, and 
the anchor bolts tightened. See the author's Machinery 
Foundations And Erection for further information. 

77. In Handling Small Turbine -driven Sets (sizes up to 
about 100 kw.) which are usually shipped completely assem- 
bled, no unusual amount of care is necessary. In general 
they can be rolled on skids without special regard to deflecting 


the bedplate; or, they can be picked up by a crane with almost 
any convenient hitch without fear of undue buckling. They 
are thus readily set on the foundation where the bedplate 
may be leveled by means of supporting wedges (Sec. 76) 
although very careful leveling is not usually necessary. 

78. In Handling Medium-sized Turbine -driven Sets (150 
to 2,000 kw.) the bedplates should be given uniform support 
to insure against deflection (bending) by the heavy supported 
parts. These machines are usually shipped assembled except 
for small parts. If rolled on the skids, great care should be 
exercised to see that the skids are supported at a number of 
points. Likewise, when hitched to a crane hook, the heavy 
parts should be supported individually instead of being carried 
on an unsupported part of the bedplate. When set on the 
wedges on the foundation top, the bedplates should be very 
carefully leveled by means of the spots or surfaces provided 
therefor; see the author's Machinery Foundations and 

Note. — A Medium-sized Turbine-driven Unit May Be Aligned 
At Its Coupling; see Sec. 167 for the method. When so ahgned, how- 
ever, account must be taken of the fact that, after the turbine-end of the 
unit is heated by the steam which it contains when operating, it will have 
expanded and will stand at a higher elevation than when cold. Allow- 
ance must be made for the amount that the turbine end will rise; see 
Sec. 85. 

79. In Erecting Large Turbo-generator Sets it is important 
to plan the work as completely as possible so that the erection 
will progress smoothly and that the man in charge can give 
his entire attention to the work at hand without fear of getting 
*'hung up" or wasting time and labor. The following pro- 
cedure (based on an article by E. H. Thompson in Power, 
July 6, 1920) will be of value in such work: 

1. When The Machine Is Received on freight cars, the various 
parts must be identified and arrangements made for unloading. It is 
necessary to consider which parts are first needed, which are to be stored 
until later, where and how these are to be stored, and how transported 
to the foundation when needed. In most plants the cars are brought 
within reach of the power-house crane; often it is necessary to roll or 


drag them to the crane. Sometimes the plant may be in the process of 
building with no crane in operation, then rigging work is the largest 

2. The Shipping Lists Can Be Checked as the unloading proceeds. 
Meanwhile a shack can, if necessary, be erected near the installation for 
tools, storing delicate parts, blueprints, and for the convenience of the 
men. Wedges and blocks for the grouting and special tools can be 
ordered for the work, to be ready when needed. 

3. The Bedplate Is First Placed On The Foundation. — A sec- 
tional baseplate should be assembled by either heating the bolts or driv- 
ing the wrench with a sledge. The bedplate should then be checked for 
accuracy. The bedplate may then be located on the foundation accord- 
ing to the center lines shown on the plant-design drawing. All openings 
in the foundation — for pipe connections, generator air ducts, drains and 
the like — should be checked for accuracy. Sometimes it is well to check 
openings and connections by temporarily assembling parts of the turbine 
casing or generator. A little such forethought may obviate the necessity 
of moving a 100-ton condenser or of chipping a concrete opening at the 
last minute, or of straining pipe flanges to make connections and causing 
a bad joint, or other trouble. The bedplate can then be carefully leveled. 

4. The Bearing Pedestals, Turbine Casings, Generator And 
Other Parts Which Must Be Aligned may now be placed on the bed- 
plate. A steel wire is generally used for aligning. The end bearings 
are first carefully doweled and bolted into their permanent position. A 
new steel wire 0.008 to 0.010 in. in diameter, such as piano wire, is tested 
to breaking strain by lifting various weights with it. The line is then 
stretched between two rigid supports, such as heavy timbers or con- 
venient columns or pieces of machinery and a tension is produced in it 
by suspending from it a weight of about ^i of its breaking load. 

The Line Is Now Moved Up Or Down Or Crossw^ise at each end 
until it is exactly central with the bored surfaces of the end-bearing 
pedestals or other parts used as a permanent guide. Wedges to suit the 
rigging are convenient in making small changes in position of the wire. 
The distance from the wire to the bored surface can be roughly measured 
with an inside caliper and with final accuracy by an inside micrometer 
or pin gage. The pin gage is generally the machinist's choice and is 
made by selecting a piece of wood }i to ^ in. in diameter, and ^ in. 
shorter than the average measurement. A pin or needle is driven in at 
each end so that measuring is done between the two pinheads or needles. 
The distance is changed by driving the pins in or pulling them out. The 
position of the wire must be adjusted so that the radial distance to the 
bored surface is the same at each side as well as above and below. It is 
not difficult to obtain an accuracy of 0.000,5 in. 

After The Tight Line Has Been Set, the other bearing pedestals, 
turbine casings, generator, gear housing, etc., can be adjusted so as to be 
central with the line. They should then be doweled, by drilling and 
reaming and accurately fitting dowels. 


The Middle Bearing Will Be Lower Than The Outside Bear- 
ings by an amount equal to the sag in the wire. This would ordinarily 
be negligible up to 15-ft. span of wire. It can be checked by assembling 
the rotary element and opening up the coupling; see Sec. 167. The mid- 
dle bearing should be shimmed until the distance between the flanges of 
the coupling is the same above as it is below, or perhaps 0.002 to 0.006 
in. more on top. 

The Turbine Casing Is Usually Set Low to allow for expansion. It 
is best, if any question arises in this connection, to get the manufac- 
turer's information on this point. It can readily be settled, however, 
by making the adjustment that seems best, and then checking the align- 
ment with shaft again, when the machine has been put in service, and 
correcting to be central under working conditions of temperature. If 
the shaft is sprung, the amount of the "spring" must be measured, and 
this should be taken into consideration when centralizing with the shaft. 

5. The Grout Should Be Poured under the bedplate after the fore- 
going questions of checking location, alignment, pipe connections, 
etc., have been settled. While this is hardening, the parts next to be 
assembled should be cleaned and made ready. 

The Importance Of Cleaning is seldom realized by the inexperienced. 
The practical man knows that grit or sand in a bearing running at high 
speed can cause considerable damage in a few minutes. A scraped joint, 
where no gasket is used, is diflScult or impossible to make steam-tight 
when dirt is present. Dirt causes alignment troubles when shims and 
contact surfaces are not properly cleaned. Cleaning is something of an 
art. An appreciable skill is required in preparing, with a sharp machin- 
ist's scraper, a scraped surface which has been warped by bolt pressure 
and expansion, and which is covered with sticky dope. Some judgment 
is required in getting this work done by unskilled and unreliable labor. 
When large pieces are cleaned, it is a good plan to provide putty knives 
or old files ground to a dull edge. The work can then be inspected by an 
experienced man, and filing or scraping done as required. 

Delicate Parts Which Are Being Assembled Should Be Covered 
With A Tarpaulin Or Other Shield each night to prevent dust settHng 
from the air, and solid particles, such as bolts or nuts, finding their way 
in. The steam passages must be continually guarded and inspected. 

The Most Difficult Part of The Assembling Is Usually The 
Turbine Rotor And Casing. It is most important to have reliable 
men to watch different parts as the lowering is done. 

The Remaining Parts To Be Assembled, which include generator, 
packing casings, steam chest, valve gear piping, etc., often require much 
painstaking work and represent a large part of the job. 

80. Casings Of Long Horizontal Turbines Are Usually Bolted 
Down At One End Only. Due to the difference in length of a 
long casing when hot and when cold, it is necessary that one 


end be allowed to slide freely. The General Electric Co. gives 
the following directions in connection with the installation of 
their 12-stage Curtis turbines: The bolts holding down the 
standard at the high-pressure end of the machine should not 
be drawn up so tight as to prevent relative movement of stand- 
ard and base at this point. The turbine casing is doweled to 
the base at a point approximately near the center of the 
exhaust passage, and expansion due to temperature changes will 
cause a movement of the standard relative to the base. Align- 
ment is preserved by keys. Marks should be placed on both 
standard and base to see that this movement actually takes 

81. To Compensate For Expansion And Wear Of Bearings, 
shims which are provided for the purpose by the manufacturers 
should be placed under the bearing pedestals. No shims 
should be used between the turbine casing feet and the support- 
ing pads of the bedplate. Insulating shims are sometimes 
necessary under the generator end bearing; see Sec. 200. 
Tests are made at the factory to determine if these shims 
are necessary. If so, they are always furnished with the 
machine. The bearings are aligned by means of a tight line 
stretched through the assembled shells of the bearings. The 
turbine end bearing must be aligned with special accuracy 
because the worm gear drive for the governor will not operate 
satisfactorily if there is any misalignment at this point. 

82. When A Turbine Is Shipped Entirely Disassembled, 
the bearings may be aligned by means of a fine steel wire tightly 
stretched through the bearing center line as explained in Sec. 
79. When the bearings themselves are received disassembled, 
they should be examined and flushed out with kerosene before 
assembling. They should, after the primary assembly and 
alignment have been made, be filled with the proper grade of 
oil. The cooling coils of the oil systems should be inspected 
for leaks by applying the full water pressure before the oil 
system is filled. Leakage of water into the oil causes much 
trouble. The oil system should then be cleaned if necessary, 
filled and examined for leaks; see also Sec. 204. 

83. The Axial Blade Clearance Of Turbines May Sometimes 
Be Tested By Means Of A Taper Gage. A plug hole, H, Fig. 


89 is usually provided for such testing in each pressure stage 
of most impulse turbines. To make the test, remove the 
plug and insert the taper gage on each side of the fixed blades 
as shown in Fig. 89. The clearance should be the same on 
each side of the fixed blades. If it is not, the difference should 
be corrected by adjusting the thrust bearing as explained in 
Div. 5. If the condition is different in two or more stages, 
the adjustment should be made so that the least clearance in 
any stage will be as large as possible. If there is no plug 
hole for measuring the axial clearance, the adjustment may be 
made by adjusting the rotor first to one extreme position and 

Fixed Blades 

Fig. 89. — Method of checking the axial clearances of a Curtis turbine. 

then the other. By carefully turning the rotor and listening 
for sounds of interference the extreme positions may be 
ascertained but this must be done very cautiously to avoid 
damage. The extreme positions may be marked on the shaft. 
See Div. 5 for further instruction. The axial alignment of 
some bucket-wheel turbines may be tested as shown in Fig. 60 
by removing the plug provided for the purpose and observing 
the alignment. 

84. Some Miscellaneous Precautions Which Should Be 
Taken In The Piping Of Large- Or Medium-sized Turbines 
are as follows: 

Piping to the turbine should be as short as possible, should be of ample 
size to prevent excessive pressure drop, should be formed in smooth 
bends whenever possible, should be so shaped that expansion will not 
strain it, and so supported that it will not bear heavily on the turbine 
casing. Cut-outs or stop valves should be provided in the branch lead- 
ing from the main header to the turbine so that the whole pipe will not 
fill with water by condensation when the turbine unit is idle. Separators 


should always be provided, where saturated steam is to be used, in the 
piping just before the steam is admitted to the turbine. Where super- 
heated steam is to be used, the use of a separator is unnecessary — 
provided that the superheat is not lost by radiation in the piping and pro- 
vided also that precautions are taken to prevent the flow of condensed 
steam into the turbine when starting. Before the piping is connected 
to a turbine, the live steam should be blown through it to remove dirt 
and scale. 

Strainers must always be used on high-pressure turbines and should 
be removable for cleaning. Strainers are usually provided by the tur- 
bine manufacturer just ahead of the governor valve. If none is provided, 
one should be procured and inserted. For low-pressure turbines using 
steam through a separator from a reciprocating engine, the strainer is 
sometimes omitted. 

Drains should be provided to take the drips from the throttle valve, 
separator, and exhaust end of the turbine casing, and low points in the 
piping where water is likely to collect. These drains may usually be 
combined and run to the condenser. A valve must be provided at the 
head of each drain to close it off as soon as all the water is removed. 
Where the condenser is located too high to take the drains, a trap should 
be provided which will deliver the drips to the hot well. 

Casings Should Be Protected From Piping Strains and all other 
kinds of strains. The capslab (supporting slab) of the foundation should 
be so rigid that no deformation is possible. The grouting of the bed- 
plate to the slab should be so thorough that no uneven support is formed 
which will cause warping. The condenser connections and other low- 
pressure steam connections should (unless a spring-supported condenser 
is provided) be made with expansion joints so that no strain will be trans- 
mitted from the condenser or other structure to the casing; see the 
author's Machinery Foundations And Erection. The relief valves 
on the casing and the atmospheric relief valve on the condenser connec- 
tion should be in good condition to avoid straining the casing or shutting 
down of the unit in case of a condenser failure. The relief valve should 
be set for about 2 lb. per sq. in. gage. There should be no hand-oper- 
ated valve which can prevent the steam escaping through the relief valve. 

85. The Final Alignment Of Turbine -driven Units On Their 
Bedplates Or Soleplates Is Preferably Made When The 
Unit Is At Operating Temperature. — The steam end of the 
unit expands when heated and, if aligned while cold, will not run 
true unless allowance is made for the expansion. The steam 
end of the unit should, for condensing operation, when cold, 
be lower than the generator end by about 0.005 in. per ft. of 
vertical distance from the point where the casing is supported 
to the shaft center and 0.01 in. lower for non-condensing 


operation. As turbine-driven boiler-feed pumps are almost 
as hot on one end as the other, very little allowance need be 

made for their expansion. 

Pellef Valye^. 


Casing \ 




Fig. 90. — Small Westinghouse tur- 
bine so supported that changes in 
temperature will have little effect on 
the shaft alignment. 

Note. — Some Small-turbine Cas- 
ings Are Supported At The Height 
Of Their Shafts (Fig. 90) so that 
increased or decreased temperature 
will have little or no effect on the 

86. The Governor And Its Oper- 
ating Mechanism Should Be In- 
spected to make sure that it has 
not been damaged in shipment. 
If necessary, the governor should 
then be adjusted to a prehminary 
setting (see Div. 6). If a gener- 
ating unit which is being installed is to operate in parallel 
with units already installed, the governor should be carefully 
adjusted for the same speed regulation as the other units (see 
Div. 6 and the author's Steam-engine Principles And 
Practice). The amount of speed variation obtainable with 
the synchronizing mechanism should be noted and adjusted 
if necessarj^ to permit synchronizing. 

87. Instructions For Checking Alignment, are given by the 
General Electric Co. as follows for their one- and three-stage 
turbine alternators : Bosses for checking the axial alignment in a 
horizontal plane will be found on or close to the horizontal center- 
lines of the unit. One of the bosses is located on the wheel 
casing, close to the outboard end. A second boss is located on 
the connection piece adjacent to the stator. A third boss is 
located on the generator outboard-bearing bracket. The two 
outer bosses are tapped for studs to carry a tight line. A 
0.016,6-in. piano wire weighted by a 30-lb. weight and supported 
at its weighted end by a small roller carried between collars 
is stretched along the unit. The horizontal distances from the 
tight line to the bosses on the connection-piece, wheel case 
and generator bearing-bracket are stamped on these bosses. 


To check this alignment, it is necessary to duplicate the dis- 
tances to the outer bosses and to compare the observed dis- 
tance to the middle boss with the figures stamped on it. No 
correction for sag of wire is necessary, but alignment must 
be checked when the unit is cold. 

Note. — The Alignment In A Vertical Plane Is Checked With 
The Same Wire stretched in the same way across the proper bosses. 
One boss for this alignment will be found on the outboard end of the 
wheel casing, between the two bosses used for leveling; a second boss is 
located on the generator end bearing bracket. One of the two bosses on 
the connection piece used for the leveling serves as the third boss. 


1. Give in order the various steps in installing a turbine. 

2. Why must the foundation of a large turbine be completely planned' 

3. What is the general form of foundation most frequently used for large condensing 
turbines? For small non-condensing turbines? 

4. Name several things which should be done in unpacking a turbine which is received 
for immediate installation. 

6. How high above a foundation should a turbine bedplate be supported for grouting? 
How soon after pouring should the excess grout be trimmed off? 

6. Why are long horizontal turbine casings securely bolted down at one end only? 
How is the alignment of the other end preserved? 

7. How may the cooling coils of an oil system be inspected for leaks? 

8. Explain a method of testing the axial blade clearance of a turbine. 

9. When is it unnecessary to provide a strainer for a steam turbine? When is a 
separator unnecessary? 

10. Where should drain pipes be provided in and around a steam turbine? Where 
should the drains lead to? 

11. How does an atmospheric relief valve protect a turbine casing? 

12. Why should a turbine preferably be finally aligned at operating temperature? 




88. The Satisfactory Operation Of A Steam Turbine 
Depends Largely On The Condition Of The Shaft, Bearings And 
Packing Glands. The operator is not particularly concerned 
with the turbine shaft except insofar as correct alignment 
(Div. 7) is concerned. This is because the shaft is designed 
and made by the manufacturer. It, if properly designed and 
made, requires practically no maintenance, except for main- 
taining proper alignment, and is not subject to operating 
difficulties. Consequently, only the more usual types of shafts 
are briefly described in Sec. 89. Bearings and the packing 
glands however, may require considerable attention and 
maintenance on the part of the operator if the turbine is to 
operate satisfactorily, the bearings and the glands must be 
kept by him in the best possible condition. 

89. Turbine Shafts, Which Represent Typical Construction 
as employed by different manufacturers, are shown in Figs. 

on Throweri. . r, • / ^ • 

Thrusr Rings- 

'•'••Governor location Coupling Key way' ' 

Fig. 91. — ^Shaft of a De Laval turbine showing key ways for fastening the discs and 


91, 92 and 93. The shafts of impulse turbines (Fig. 91) 
are nearlj^ always made solid, while those of reaction turbines 
(Figs. 92 and 93) are generally hollow. The shafts of practi- 
cally all turbines are now made ''stiff." See note below. 

Note. — Some Manufacturers Apply The Term "Spindle" to 
designate the complete rotating element, as in Figs. 92 and 93. How- 
ever, the terms "shaft" and "spindle" are generally synonymous. 


Sec. 90] 



Note. — The "Critical Speed" Of A Shaft which carries a load, as 
for instance a turbine rotor, is the specific speed at which the shaft 
vibrates most violently. If the shaft is permitted to rotate for any 
length of time at its critical speed, the vibrations may prove disastrous. 
The explanation for this vibration is too technical to be given here. It 

LP. Balance Piston^ LP. Spindle 

\LP.O/ISIin^ \ iU'l—UHaJt^re*^ V '^ X 


^ H.P.Spindle 
H.P. Gland Runner-' ^^^ Thrust) 

Col la r--' 

^nd / 1 ^ ^a, .1. . ' "M ua^ ^ 

LP. GlancI Runner I.P.Spindle Rings'- 

Fig. 92. — Section through the spindle of a Allis-Chalmers reaction turbine. (L.P,. = 
low pressure. I. P. = intermediate pressure. H.P. = high pressure.) See also Fig. 

is a fact, however, that at speeds well above or well below their critical 
speeds, all shafts (unless badly unbalanced) will run fairly free from 
vibration. In the early days of steam-turbine engineering most tur- 
bines were operated above their shafts' critical speeds. In starting or 
stopping such turbines it was essential that the critical speed be passed 
as quickly as possible. Nowadays, however, nearly all turbine shafts are 

Inf^- mediate P-essure 

- «%;;:^?s: 




^\ Jet A 








Spina ir 

■X," '^ 


^^V "^'''"'9 


Of/ Sling' 

; '-Labyrir. 



\' ■ 

High Pressure balance P/sfon 

Fig. 93. — Rotor or spindle of a reaction turbine. {Allis-Chalmers Mfg. Co.) 

designed to be so "stiff" that the turbines operate normally at speeds 
well below the critical speeds of their shafts. Furthermore, the rotors 
of modern turbines (the better ones) are carefully balanced in the 
manufacturers' shops to further lessen the dangers due to vibration. 

Note. — A Turbine Shaft Is Said To Be "Stiff" If It Is Designed 
To Operate At Some Speed Below Its Critical Speed. A turbine 
shaft is said to be "flexible" if it is designed to operate above its critical 

90. The Two Principal Types Of Bearings In Steam Tur- 
bines are: (1) The main bearings, which carry the weight of 


the rotor and which also prevent any excessive movement of 
the rotor in any direction perpendicular to its axis; the main 
bearings are sometimes called radial pressure hearings. (2) 
The thrust hearings which restrain the rotor from excessive 
movement in either direction parallel to the axis of the rotor. 
Bearings of each of these types are discussed in the following 

91. Table Showing Classification Of Steam-turbine Bear- 
ings. (Only those bearings are included in this table which 
restrain the movement of the rotor. The bearings which are 
used in connection with the governor, the oil pump, or other 
subsidiary apparatus are not included in this table and are 
not discussed in this division.) 


Main bearings 

Thrust bearings 

Plain bearings 

Ball bearings 

Flexible, Fig. 94 

Rigid, Fig. 100 

One row. Fig. 50 

Two rows. Fig. 65 

Roller, Fig. 103 

Simple collar. Fig. 104 

Multi-collar, or marine, Fig. 105 

Ball, Fig. 106 

Kingsbury, Fig. 108 

92. Plain Flexible Steam-turbine Bearings (Table 91 and 
Fig. 94; see also the note below) generally consist of: (1) 
The hahhitt, B, which contains the oil grooves, G, and upon 
which the journal bears. (2) The lining, L, which is held in 
place by the spherical seat, S. The Hning is usually split 
along the horizontal center line; thus, it is divided into two 
parts, which are sometimes called the upper lining and the 
lower lining. (3) The pedestal, P^ which supports the lining 

Sec. 92] 



through the seat. (4) The bearing cover, C. Various manu- 
facturers employ different constructional details in flexible 



I-Longitucrf noi ( Section IL-Transverse Section 

Fig. 94. — Spherical-seated bearing of Allis-Chalmers steam turbine. 

bearings. This is evident from a comparison of Figs. 94, 95, 
96, 97, and 98. Flexible bearings of some kind are used in 

I-Longi+udinal 5ec+ion II-Transv«rsal Section 

Fig. 95. — Spherical-seated steam-turbine main-bearing. (Oil enters at D and passes 
upward through the spaces E, entering the bearing through the groove at F. The bab- 
bit is so bored that the horizontal "diameter" dimension is slightly greater than the 
vertical "diameter" dimension. (Westinghouse Electric & Mfg. Co.) 

nearly all steam turbines. Bearings of this type are also 
called spherical-seated hearings, and self-adjusting hearings. 
The function of a flexible bearing is explained in the following 


Note. — A Flexible Bearing So Operates that the bearing will, 
ivithout causing excessive friction, automatically adjust itself so that 

$,^hf Hole Plug 

■Lining Screw 

Fia. 96. — Section through outboard bearing of General Electric Co. steam turbine. 

.. ,, .... /" Spherical Seat 

r'"" upper 








-Oil Passageway 
X-Transverse Section JT-Lpngitudinal Section 

Fig. 97. — Section through bearings of a Kerr turbo-generator. 

the axis of the bearing coincides with or remains parallel to the axis of 
the journal or shaft. See Fig. 99. The axis of the shaft when in its 
normal position is indicated by the center line, A. If from any cause, 




Casf-Iron ^ 

Lining "^;, 

l-Tubes Assembled 

E-A55emb\ed \n Cas-t-lron 

EL-Showing Relative Posi + ions 
of Tubes 


^ , - -Re fain in g Nuf 

BC-lnnermosi Bronze Tube 

Fig. 98. — Self-adjustable or flexible main bearing consisting of a nest of tubes for 
high-speed turbines of small capacity. [The bronze tubes, E, D, and C (III) fit over 
each other (II) with some clearance, so that the innermost is free to move slightly in any 
direction. Oil fills the clearance between the tubes and forms a cushion which tends 
to dampen vibration.] 


Shaft Ands^ 

Bearing Cover 



Fig. 99. — Illustrating action of a spherical-seated "flexible" bearing, when the turbine 
shaft bends. The bending is exaggerated for purpose of illustration. 


the shaft should bend, its axis will then be at some other position as 
indicated by B. Now, since the bearing lining is held in a spherical 
seat {S, Fig. 94), the bearing will rotate in this spherical seat and assume 
the position shown in Fig. 99; thus, coinciding with the axis of the shaft. 
If the bearing were a rigid hearing (Fig. 100) the bearing could not readily 
adjust itself to any bending of the turbine shaft. Consequently, if the 
turbine shaft should bend, excessive friction would result and the bearing 
would be subjected to excessive wear, and probably to heating. 

.■Threaded Collars Locked In PosWon rprm A Thrusf Bearmgr-^ 

Fig. 100. — Rigidly seated steam-turbine bearing. (Any bending or deflection of the 
shaft will tend to distort the housing and pedestal, thus causing excessive bearing fric- 

93. Ball Bearings Are Used As The Main Bearings Of 
Small-capacity Turbines by some manufacturers. A double- 
race, self-adjusting ball bearing is shown in Fig. 65. The 
advantages claimed for main bearings of this type are that 
they minimize friction and are readily accessible for renewal. 
Ball bearings are seldom used for turbines of capacities greater 
than about 200 hp. A ball bearing should be flushed out 
occasionally with kerosene. A ball bearing cannot be 
repaired; if it becomes badly worn, it must be renewed. 

Note. — The Relative Location And The Constructional 
Arrangement Of Steam-turbine Main Bearings are shown in varioua 
illustrations in Div. 3. 

94. In General, The Temperature Of The Oil Leaving A 
Turbine Main Bearing should not exceed about 150° F. 

Sec. 95] 



See also Div. 10. However, there are some turbines the 
bearings of which are designed to operate at a temperature of 
from 195° to 212° F. To prevent excessively-high bearing 
temperatures, the main bearings of some medium- and large- 
capacity turbines are (Fig. 101) water cooled. Cold water is 
forced through the coils which are imbedded in the bearing 
lining. Those turbines which have circulation lubricating 
systems (Div. 10) are generally equipped with separate 


.'Spherical Seat 

-Bearlnof Coyer 

Upper Half Lining 

I-Partial Longi+udinal 

doffpm / H-Sec+iion X-X 

Half Linincf j v, . v^. - . . vv|. . . . ^ . ^ 

(End View) 

' Retaining Clip 
\ for Pipe Co! t 

Cooling Coil- -■■'^, tj 

IT-Cylinolrlcal Sec+ion Of Bo++om 
Waif of Lining Showing Cooling Coil 
Fig. 101. — Showing constructional arrangement of a water-cooled steam-turbine bear- 
ing. (General Electric Co.) 

coolers for lowering the temperature of the oil after it has 
passed through the bearing. 

95. The Care Of The Main Bearings Of A Turbine consists 
principally in providing proper lubrication (see Div. 10). If 
proper lubrication is not maintained, excessive wear of the 
bearing will result, or the bearing may be burned out. Exces- 
sive wear in the bearing will disturb the alignment. This will 
usually cause undue vibration which will, in turn, cause the 
bearing to wear still more. If a slight misalignment due to 
wear is discovered in time it may be corrected by removing 
and inserting shims (Fig. 94) which are generally provided 
between the lining and the blocks which support the lining. 
With proper care, a turbine main bearing should last from 6 
to 10 years. 


Note. — Excessive Wear In The Bottom Of The Bearing Lining 
Usually Results In One Or More Of The Following Conditions: 

(1) Misalignment of the hearings. (2) Shoulders along the oil groove, 
which will cut off lubrication and cause heating. (3) Contact between 
some stationary and some rotating -part of the turbine. (4) Hard parts of 
the babbitt wearing the journal irregularly. (5) Excessive clearance in the 
upper half of the liner, which may permit the oil to be thrown out. Obvi- 
ously, the remedy is to install a new lining or to rebabbitt the old lining 
(Sec. 97). 

96. A Turbine Bearing May Be Repaired by: (1) Installing 
a new bearing which has been supplied by the manufacturer. 

(2) Rebabbitting the old bearing. If a reserve bearing is at 
hand, the first method is the preferable one. This is because 
that, by using the reserve bearing, the necessary repair can be 
effected with a minimum loss of time, and also because a better 
fit will probably be secured. However, if an extra bearing 
is not readily available, new babbitt can be poured into the old 
lining and good results will obtain if the work is properly done. 
In any case, if a bearing is destroyed by any means except by 
ordinary wear, the cause of the destruction should be located 
and removed before the turbine is again put into service. 
A method of rebabbitting a turbine bearing is described in the 
following section. See also the author's Steam-engine 
Principles and Practice. 

97. In Rebabbitting A Turbine Main Bearing, the original 
dimensions and shape of the old bearing should, if known (Fig. 
97), be followed as closely as possible. If the original dimen- 
sions of the old bearing are unknown the new bearing can be 
made as explained below. 

Explanation. — Pour the babbitt so that the diameter of the bearing 
is the same as that of the journal. Then scrape out the oil grooves. 
The oil grooves should (Fig. 102) be about }-i in. deep and about % 
in. wide. Two straight grooves (Fig. 102) are all that are generally required. 
Some turbine bearings have only one oil groove (Fig. 95) which is located 
in the top of the bearing. The location of the grooves will, for a forced- 
circulation lubricating system, be determined by the holes in the cast- 
iron shell for the oil inlet and outlet (Figs. 94, 96, 97, and 102). To pre- 
vent excessive oil leakage from the bearing, the ends of the groove should 
be about K in. (Fig. 97) from the ends of the bearing. The square edges 
of the groove should be scraped away to a rounding contour (Fig. 102) 
so that there will be no sharp edge to interfere with the oil film. 

Sec. 98] 



After making the grooves, the next step is to fit the bearing to the 
journal. First put the lower half of the bearing in place. Then scrape 
out this lower half (see the author's Steam-engine Principles And 
Practice) so that for about 55 deg. (Fig. 102) from each side of the verti- 
cal center line the bearing is an exact fit for the journal. Be sure to 
remove all high spots from this portion of the bearing. From the 
extremities of this area — that area which is fitted to the journal — up to 
the lower edges of the grooves, the bearing should be scraped away 
slightly {A and A, Fig. 102) so that a wedge-shaped oil-film space will 
be provided. A clearance should be provided between the journal and 
the upper half of the bearing. This clearance should be about 0.002 in. 
for each inch of journal diameter. That is, for a 2-in. journal, the clear- 
ance should be about: 2 X 0.002 = 0.004 in. This clearance may be 

■Vertical t 


j Direction Of: 
' 5haff Rofafion 

Fig. 102. — Illustrating one method of re-babbitting the main bearing of a turbine. 

obtained by inserting shims {S and S, Fig. 102) of the proper thickness 
between the upper and lower halves of the lining, and then scraping 
away the bearing at B and B. If this clearance is too small the oil pas- 
sage will be restricted. If it is too great, there may be an excessive oil 
leakage. The clearance above the journal can be determined by putting 
a piece of soft lead fuse wire on the top of the journal and then tightly 
bolting on the upper half of the lining. Then remove the upper half of 
the lining and caliper the mashed fuse wire. 

98. The Primary Function Of Steam-turbine Thrust 
Bearings is to hold the shaft in such an axial position that 
proper clearance will be maintained between the rotating 
and stationary parts. Since impulse turbines are inherently 
subjected to but little end thrust and reaction turbines are 
generally provided with dummy pistons (Sec. 67) for balanc- 
ing the end trust, the thrust bearings are not usually (see 


Sec. 99) required to withstand much pressure. However, 
where a governor or an oil or water circulating pump is driven 
from the turbine shaft through a heUcal gear, a considerable 
end thrust may be exerted. Thrust bearings of the principal 
types, also some methods of adjusting them, are described in 
following sections. 

99. The Four Principal Types Of Thrust Bearings which are 
used in steam turbines by the various manufacturers are: (1) 

Bolt, For 
Thrust Cage 
And Adjusting 
Position Of 



Thrust Cage .' 

Moved Axial ly ' 

By Adjusting Bolt, A, 

And Prevented From 

Turning By Key, K. 



Fig. 103. — Roller thrust bearing. (The hardened steel washers, S, are held stationary 
by the dowels. The washers, M, rotate with the shaft. The rollers, R, roll between 
M and S. Clearance between R, S and M is adjusted by removing B and turning C. 
Axial position of shaft is adjusted by A.) 

Roller thrust hearing, Fig. 103. (2) Collar thrust bearing, which 
may consist of only one collar, Fig. 104, or of a number of 
collars, Fig. 105. (3) Ball thrust bearing, Fig. 106. (4) Kings- 
bury thrust bearing. Figs. 107 and 108. The operation of the 
bearings of the first three types will be evident from a study of 
the respective illustrations. The operation of the Kingsbury 
thrust bearing is explained below. 

Explanation. — The Kingsbury Thrust Bearing is sometimes (Fig. 
108 and Fig. 69) contained within the main bearing lining. Sometimes 
it is mounted in a separate casing on the end of the shaft, as in Fig. 107. 

Sec. 99] 



This bearing is arranged to withstand thrust in the direction of arrow A 
against the bearing blocks, F (Fig. 108-/). One block, G, is placed on 

.-Main Bearing Lining 
Bearing Cap-^ n^i Thrust Rlncf^ 

'Vrah Plugs-'' 

Fig. 104. — Simple collar thrust bearing. The two thrust rings R and R are pinned 
to the oil deflectors, Z), and rotate with the shaft. Axial movement of the shaft is 
restrained by contact of these rings with the ends of the lining of the main bearing which 
are faced with babbitt. (General Electric Co.) 

Graduated Dial On Upper 
Adjusting Screw 

Fig. 105. — Multi-collar or marine-type thrust bearing. {Westinghoitse Electric & Mfg. 


the side opposite from the direction of thrust to restrain any endwise 
movement of the shaft. The bearing blocks, F, (as shown in ///), rest 


against pivoted projections on the equalizing blocks, K. The purpose 
of the equalizing blocks, K, is to equalize the pressure of each bearing 

fnspecHon Bal I -Thrust 
Cap-'^ '.Bearings 

Fig. 106. — Showing the use of the Gurney Ball Thrust Bearing for maintaining axial 
shaft alignment. {Terry condensing turbine.) 

Sfeam-End 5ieam-End Bearing 


Case Cap 

■Thrust Collar 

Thrust Block 
Bearinef Adjusting Screyi 

Fig. 107. — Kingsbury thrust bearing parts of Moore steam turbine (Instruction Card 
No. 3). Axial adjustment of the rotor is made by turning screw -S, which moves a slide 
r, carrying the thrust bearing with it. The thrust blocks of this Kingsbury thrust 
bearing are also adjustable so the correct clearance between the blocks and the collar 
can be obtained. A clearance of about 0.004 to 0.005 in. on each side or 0.008 m. to 
0.010 in. total, is recommended. 

block on the collar, E. The collar, E, is shrunk on the shaft and rotates 
with it. 

Sec. 100] 



Thus if any portion of E (Fig. 108-7//) tends to exert a greater pressure 
on some one bearing block, say block F, the equalizing block, K, on which 
this particular bearing block is pivoted is pressed downward. This 
causes the two adjacent equalizing blocks Ki and Ki to rotate a little, 
which causes the next equalizing blocks K2 to push upward on the next 
bearing block, F2. Thus, the total thrust which is exerted by E in the 
direction of A (Fig. 108) is always equally divided between all of the 
bearing blocks. Also, the total thrust on any one bearing block is uni- 
formly distributed over the face of that block. Consequently every 



I-Longitudlina\ Sec+ion 
Bearing Block- 

TI-Transverse Section A-A Y<i+h 
Shaft And CoWar Omitted 

Hi-Cylindrical Section 
Fig. lOS. — Kingsbury thrust bearing. 

minute portion of the face of each bearing block is always active in carry- 
ing the thrust. This design and construction produces (automatically) 
a wedge-shaped oil film at L (Fig. 108-1 1 1), which provides effective 
lubrication at all times. 

Inasmuch as a thrust bearing of this type is capable of satisfactory 
operation under very high unit pressures (350 to 500 lb. per sq. in.), the 
area of the balance pistons of reaction turbines is sometimes reduced 
and the Kingsbury thrust bearing is designed to carry the unbalanced 
end thrust. To insure that the end thrust will always be against the 
bearing blocks, turbines are {Westinghouse Electric & Mfg. Co.y 
"Instruction Book No. 5,171"), sometimes installed with the thrust- 
bearing end lower than the other end by about 0.02 in. per foot of length 
of the turbine. 

100. The Axial Adjustment Of A Turbine Rotor Determines 
The Axial Clearance Between The Rotating And The Sta- 


tionary Parts Of The Turbine Proper. If the axial position 
of the rotor is not properly adjusted, the turbine will not 
operate at its maximum efficiency; and in case of extreme axial 
mal-adjustment, the turbine may be wrecked. As suggested 
in Sec. 98, the axial adjustment of the rotor is generally made 
by shifting the thrust bearing. Various methods of making 
this adjustment and the amount of clearance which is neces- 
sary between stationary and moving parts of the turbine are 
discussed in following sections. 

101. The Types Of Mechanisms Which Are Generally 
Employed For Axial Adjustment Of A Turbine Rotor are: (1) 
Screws or nuts, Figs. 103, 105, and 107. (2) Shims, Fig. 104; 

Steam Jer-' 
Fig. 109. — Showing correct amount of "lap" for a Terry tangential-flow turbine. 

see also Fig. 57. Screws are used for effecting axial adjust- 
ment in practically all steam turbines. However, in turbines 
of the smaller capacities, some manufacturers use shims (Fig. 
104). The detailed design of the mechanism for axial adjust- 
ment which is used by one manufacturer differs from that 
used by every other manufacturer. Even the axial-adjust- 
ment mechanisms for turbines made by the same manufacturer, 
but of different types and capacities, are different one from 
the other. Consequently, it is impractical to treat herein the 
various mechanisms which are employed for this purpose. 
The operator should, by a careful study of the machine and of 
the manufacturer's instructions, thoroughly familiarize himself 
with the adjusting mechanism before attempting to make an 
axial adjustment of the rotor. 


102. The Axial Adjustment Of A Tangential-flow Turbine 

(Sec. 56) is usually made by providing the proper ''lap" 
between the wheel and the reversing chamber. See Figs. 
60 and 109. 

103. The Axial Adjustment Of An Axial-flow Turbine Rotor 
Which Is Provided With Adjusting Screws is usually made 
as follows: The turbine after being heated to its operating 
temperature (Div. 11) is throttled down so that it runs at 
about 10 per cent, of its normal speed. While running at this 
decreased speed, the rotor is, by the axial adjusting mechanism, 
moved in a longitudinal direction until a slight rubbing is 
heard. Then, the adjusting mechanism is operated in the 
opposite direction until a slight rubbing is again heard. In 
making this second movement count the number of nut or 
screw turns which are made. Now, move the rotor back in 
the direction of the first movement by one-half the number 
of nut or screw turns just counted. Next, by whatever kind 
of locking device that is provided, lock the rotor in this position. 
This should locate the moving part in the center of its minimum 
clearance, which, for most axial-flow turbines, is the correct 
axial position for the rotor. 

Note. — The Slightest Rubbing May Be Readily Heard by hold- 
ing one end of a short piece of gas pipe or a file against the casing and 
the other end near the adjuster's ear. This rubbing should not be per- 
mitted to continue longer than an instant, and should not be severe. 
Otherwise, the turbine is likely to be damaged. 

Example. — Assume that the rotor of the 3,600-r.p.m. turbine the 
thrust bearing of which is shown in Fig. 103 is to be axially adjusted. 
First heat up the turbine. Then, throttle down to about 350 or 400 
r.p.m. Turn bolt A in a right-hand direction until a rubbing is heard. 
Then, counting the number of turns, turn A back in a left-hand direction 
until a rubbing is again heard. Now turn A in the right-hand direction 
one-half the number of turns just counted. Next, lock A in this position 
with the locknut, N. 

Note. — The Axial Rotor-adjustment Of Those Turbines Which 
Have Axial-clearance Metallic Labyrinth Glands (Sec. 112) Can- 
not Be Made As Described Above. This is because the axial-clearance 
labyrinth glands (Fig. 112) must have a small axial clearance between 
the rings, R, on the balance piston and the tips, T, of the dummy rings. 
The proper value of this clearance varies with the size and design of the 
turbine and must be obtained from the turbine manufacturer. 


104. The Axial Adjustment Of A Turbine Rotor Which Is 
Provided With Adjusting Shims (Fig. 104) must be made in 
a manner somewhat different from that which is described in 
Sec. 103. This is because the shims are not readily accessible, 
and therefore the adjustment cannot generally be made while 
the turbine is at its operating temperature. Consequently, 
the adjustment must be made while the turbine is relatively 
cold. Therefore, in effecting the adjustment while the turbine 
is cold, due allowance must be made for the expansion of the 
shaft and casing which will occur when the turbine is heated 
to its operating temperature. Usually this allowance may be 
made by adjusting the axial position of the rotor as explained 

Explanation. — Take down the bearing (Fig. 104) and fill the shim- 
spaces at A and B with shims so that one-third of the thickness of all 
shims will be in the space at B and two thirds of the thickness of all 
shims will be in the space at A. This will (Fig. 70) locate the rotating 
discs to the right of the central position. Then, when the turbine is 
heated during operation, the expansion of the shaft occurs away from 
the thrust bearing toward the exhaust end. This expansion moves the 
discs to the left (Fig. 70), and the rotating blades will take a position 
nearly central with the stationary blades and nozzles. 

105. Thrust Bearings Of The Collar, Roller, And Kingsbury 
Types Must Be Adjusted after the axial adjustment of the 
rotor is made. If the thrust bearing is too tight, it will bind 
and may burn out. If it is too loose, the correct axial position 
of the rotor will not be maintained. A thrust bearing of the 
collar, roller, or Kingsbury type should have a total clearance 
of from about 0.008 to 0.010 in. That is, the thrust bearing 
should be so adjusted that the shaft will have a ''play" in 
the axial direction of from about 0.008 to 0.010 in. This 
adjustment may be made by screws which are (Fig. 107) pro- 
vided for this purpose; or in the case of a shimmed bearing 
(Fig. 104) which has no screws, a 0.004-in. shim is, after the 
axial adjustment (Sec. 104) has been made, removed from each 
of the shim-spaces. 

Note. — Ball Thrust Bearings Usually Require No Adjustment. 
They are made with the proper amount of clearance. When they wear 
so that the clearance is excessive, they must be renewed; see Sec. 93. 

Sec. 106] 



106. A Steam-turbine Gland is a device for minimizing the 
leakage of steam or air through the clearance which must be 
provided between the rotating and stationary parts. Thus, 
where the shaft passes through the high-pressure end of the 
casing (Fig. 75) a gland must be provided to prevent an 
excessive leakage of steam out of the turbine. In a pressure- 
staged or a velocity-and-pressure-staged turbine (Fig. 70) 
the pressure on one side of a diaphragm is less than the pres- 
sure on the other side. Consequently, there is a tendency 
for the steam to leak past the diaphragm along the periphery 
of the shaft. To minimize the steam leakage at these loca- 
tions, a gland of some sort must be used. When a turbine is 
operated condensing, the steam pressure within the turbine 
casing at the exhaust end is less than atmospheric pressure. 
Therefore, to prevent air from leaking into the turbine and 
decreasing the vacuum, a gland must be provided around the 
shaft where it passes through the exhaust end of the turbine. 
The repair and adjustment of glands of various types are 
described in the following sections. 

107. There Are Four Principal T5rpes Of Steam-turbine 
Glands: (1) Metallic-packed or stvffing-box gland, Fig. 110. 
(2) The metallic-lahyrinth gland, Fig. 113. (3) The ceiitrifugal 
water-packed gland, Fig. 116. (4) The carbon-packed gland 
(Fig. 120). The construc- 
tion and maintenance of 
glands of each of these 
types are treated herein- 
after in this division. 

108. Metallic-packed Or 
Stuffing-box Glands (Fig. 
110) are stuffing boxes 
which are packed with a 
flexible metallic packing. 
Glands of this type are, 
generally, used only for 

velocity- or single-staged turbines which are designed to 
operate non-condensing at low back-pressures — not exceeding 
about 10 lb. per sq. in. — and at speeds below 3,600 r.p.m. 
Since the steam pressure in the casing of a turbine of this type 

Thrust Collan 

' Locknuf Box 

Wafer Def/ector 
■ EKhausf Felt Washer' 


Bearing . 

Fig. 110. — Section through stuffing box and 
related parts of Type-6 Sturtevant turbine. 


is about the same as the atmospheric pressure, the function of 
the glands is not so much to prevent a waste of steam as it is 
to prevent any steam which condenses on the shaft from 
ultimately finding its way into the bearing. 

109. A Metallic -packed Gland May Be Repacked as de- 
scribed below. Soft metallic packing rings with skived joints 
(Fig. Ill) should be used. The rings should be about 3-^ in. 
thick. The outer diameter of the rings should be approxi- 
mately the same as the inner diameter of the stuffing box. 
The inner diameter of the rings should be approximately the 
same as the diameter of the shaft. The number of rings 


B<' , E-5ection IH-Skivcd 

I-Elevoition A-B Or Lap 


Fig. 111. — Showing skived-jointed metal packing rings. 

required will depend upon the length of the stuffing box and 
upon the thickness of the rings which are used. After new 
packing is installed, the cap (Fig. 110) should be screwed up as 
tightly as possible with the fingers. Then, after the turbine 
has been started, the cap may be tightened a little more with 
a wrench. Allow a reasonable time for the packing to adjust 
itself before making any further adjustments. All packing 
of this type will leak somewhat when the turbine is starting 
cold but the packing becomes tighter as the turbine heats. 
If the cap is screwed up too tightly, the packing will be scorched 
and ruined. Never use a wrench to tighten the gland except 
when the turbine is running. Unless a packing should burn 
out, it is seldom necessary to install an entire new packing; 
merely add a new ring as described below. 

Sec. 110] 



Note. — The Wear In A Metallic-packed Gland Should Be 
Taken Up by tightening the stuffing-box cap and occasionally inserting 
a new ring. When a new ring is inserted, it should be placed between 
the outer and the second rings of the old packing. A slight steam leak- 
age from a metallic packed gland is permissible and helps to lubricate 
the gland. But a leak that "blows" steam should not be tolerated. 

110. Metallic-labyrinth Glands (Figs. 112, 113 and 114) are, 
as the name suggests, designed to force the steam to follow a 

S-tafionctry Ca5i'r)gr.^ 

I- Radio) l-Clearcxnce Type, 
Low-Pressure Balance Pis+or 

Small Large 


Clearance Clearance f^mofS '^"/3^ 



"•Clearance From O.OOd To 0.020' Dependiny On Size Of Unit 
H-Axial- Clearance Type, High Pressure balance Pis+oh 

12. — Double-labyrinth glands to minimize steam leakage around the balance 
pistons in a reaction turbine. {Allis-Chalmers Mfg. Co.) 

long winding path through the gland. The steam, in passing 
through each constriction in the path, is subjected to a throt- 
tling action with a consequent reduction in pressure. Thus, 
the reduction in pressure and the frictional resistance which 
are occasioned in passing through the labyrinth passageway 
permit but a small amount of steam to escape. 


111. There Are Two Types Of Metallic -labyrinth Glands : 

(1) The double labyrinth gland, Figs. 112 and 113, which con- 
sists of annular rings on the rotating element which fit into 
annular grooves in the stationary element. (2) The single 
labyrinth gland, Fig. 114, which consists of a number of 
stationary annular saw-toothed projections which fit closely 


■4 Pipe Tap For 5f earn Seal 

Do//e/s -^ 

Third Segment 

Fig. 113. — Gland of Steam Motors Company turbine. This gland is of the double- 
labyrinth type and is used in all of this company's turbines. It is suitable for any 
ordinary back pressure or vacuum. A drain. D, is provided between segments 2 and 3 
of the packing. This should be piped and the drain line led away to the atmosphere. 
Or, since only clean steam drains from it, it may be led to the feed-water tank, provided 
the tank is operated at atmospheric pressure. For location of drain, see Fig. 28. For 
condensing operation a ^^-in. steam-seal pipe, in which is inserted a valve, should be 
led to S. The sealing steam is admitted to the gland between segments 1 and 2 and 
the valve should be opened sufficiently wide, that there is just a "whiff" of steam leak- 
age visible at the bearing end of the gland. The drain connection remains in any case, 
but for condensing operation it may be fitted with a valve which should be so adjusted 
that the flow of steam through the drain will not be excessive. 

to the smooth shaft. Glands of each of these types are 
described in the following sections. 

112. The Double Labyrinth Glands (Figs. 112 and 113) 
are generally used: (1) To prevent leakage of steam over the 
balance pistons (Sec. 67 and Fig. 112) in a reaction turbine; 
as indicated in Fig. 112, balance-piston labyrinth glands may 
be of either the axial-clearance or of the radial-clearance type. 
(2) To prevent steam or air leakage around the shaft (Figs. 65 

Sec. 113] 



and 113) at the steam- or exhaust-end of an impulse turbine. 
If the turbine is to be operated non-condensing against an 
appreciable back pressure, steam leakage in minimized by the 
lab3a-inth passageway as explained in Sec. 110. If the turbine 
is to be operated condensing, leakage of air into the casing 
(along the shaft) is prevented by a steam seal, the operation 
of which is explained below. 


Plate - • 

' Sprlncf 

Explanation. — The Operation Of A Steam Seal is as follows: 
Assume that steam is admitted at about the middle of the gland {S, 
Fig. 113) at a pressure of 3 lb. per sq. in. gage. The steam will leak 
through the labyrinth passageway in both directions, part of it going 
into the turbine and part outward to the atmosphere. If steam is leak- 
ing outward to the atmosphere, it is obvious that air cannot at the same 
time leak into the turbine casing. The steam which leaks into the casing 
will have practically no effect on the vacuum, whereas air would, if per- 
mitted to leak in, tend to lower the 
vacuum considerably. The operation 
of the steam seal in a carbon-packed 
gland (Sec. 118) is essentially the same 
as is described above. 

Note. — The Advantages And Dis- 
advantages Of a Double Labyrinth 
Gland are: (1) There are no ruhhiiig 
surfaces. Therefore it is frictionless 
and consequently has a long life. (2) 
It ordinarily limits the axial end-play of 
the shaft. Hence, if rubbing should 
occur and the gland is injured, a new 
gland will usually be required. The in- 
stallation of a new gland is an extremely 
difficult and expensive procedure. 


113. The Single Labyrinth Pack- 
ing Gland (Fig. 114 and Sec. Ill) 
consists of one or more metallic 
rings (Fig. 115) which are loosely 
supported by a shoulder {S, Fig. 
114) in the packing chamber. 

Each ring is composed of three equal segments (X, F, and 
Z, Fig. 115) which are held together by a garter spring (G, 
Fig. 114). One of the segments is provided with a stop to 

Fig. 114. — Single-labyrintli-t y p e 
packing gland to prevent steam 
leakage along the shaft where it 
passes through a diaphragm. 

prevent the ring from rotating with the shaft. When first 


assembled, each ring is so machined that the tips of the saw- 
tooth projections hug the shaft, and the flange {F, Fig. 115) 
clears the shoulder, S, in the packing chamber (P, Fig. 114). 
When cold, the clearance between the segments of the ring is 
(Fig. 115) about 0.005 in. When the turbine heats under 
operating conditions, the rings expand. Thus, the clearance 
between segments closes up and forms an arch hound butt- 
joint. Also, the shaft wears off the points of the teeth until 
the flange {F, Fig. 115) on the ring rests on the rabbeted 
shoulder, S, in the packing chamber, P, Fig. 114. Thus, a 
closely fitting labyrinth gland is provided, the applications of 
which are given below. 

Note. — Single Labyrinth Packing Glands Are Used In Some 
Impulse Turbines Of The Smaller Capacities for both the steam and 


5<_' 0.005 
I-Plan View H-Section A-6 

Fig. 115. — Metal packing ring of the single-labyrinth type. 

exhaust-end glands and also for the diaphragm glands. When used in a 
turbine that is to be operated condensing, single labyrinth glands for the 
steam and the exhaust ends are steam sealed in a manner which is similar 
to the steam seal used for double labyrinth glands (Sec. 112). 

Note. — Single Labyrinth Glands May Be Tightened To Take 
Up Wear by machining out the flange seat {S, Fig. 115) of the ring, and 
then filing off the ends of the segments so that the correct end clearance 
of about 0.005 in. (Fig. 115) between segments will be provided. These 
operations should be performed with extreme care so that concentricity 
and proper end-clearance will be maintained. If the ends of the saw- 
teeth are worn so that the tips are materially widened, the grooves 
between teeth should be remachined out so that the teeth are sharp. 

114. A Centrifugal Water-packed Gland (Fig. 116) is merely 
a centrifugal-pump runner, C, which is fixed to and rotates with 



the turbine shaft. Machined in the turbine casing, or in the 
gland casing, is a chamber, B, within which the runner rotates. 
Water is admitted at the inlet, A . The runner is so designed 
that when the turbine is operating at normal speed, a water 
pressure of about 20 lb. per sq. in. gage would, if the water 

Connect /on 
Pressure Gaofe' 

CPSJ^ Drain-.' Wafer Inlet-. 
Fig. 116. — Centrifugal water-packed gland. 

'--Drain To 

were admitted at the center of the runner and no outlet were 
provided, be produced at the periphery of B, Consequently, 
if water is supplied at the periphery at a pressure of about 5 lb. 
per sq. in. gage, the pump runner holds the water in a solid 
annular ring against the periphery of the chamber, C. This 
produces a hermetic seal which entirely precludes leakage. 

Note. — Any Water Leakage From A Centrifugal Water-packed 
Gland Must Be Drained Away, If the turbine is to be operated con- 
densing, the glands must sometimes be sealed for raising the vacuum 
before the turbine is started. Obviously, during the period of starting 


the pump runner does not function. Therefore, to prevent excessive 
leakage of the sealing water while starting, single labyrinth glands are 
provided as shown at E in Fig. 116. There may also be a slight leakage 
of water while the turbine is running at full speed. To prevent any water 
which may leak outward (to the left in Fig. 116) along the shaft from 
being thrown out into the engine room, an outer gland flange, H, is pro- 
vided. To prevent water from finding its way along the shaft and into 
the bearing, the drain K is piped to an open sewer or to some other 
region where the pressure is not above atmospheric. This drain, K, 
must be kept open at all times. The inner flange (/, Fig. 116) prevents 
any water which leaks inward from being thrown against the moving 
blades. The drain, G, is piped to the exhaust pipe of the turbine. Other 
methods than those described above for sealing during starting and for 
taking care of the leakage water, will be evident from the construction 
of the turbine in which they are used. 

115. Centrifugal Water-packed Glands Cannot Be Used In 
Close Proximity To High-pressure Steam. That is, if a 
centrifugal water-packed gland were used in the high-pressure 
end of a pressure-staged or composite-staged turbine (Fig. 
73), the water in the gland would tend to vaporize. Conse- 
quently, glands of this type are generally used for only: (1) 
The exhaust end of impulse turbines, Fig. 69. (2) The steam 
and exhaust ends of low-pressure impulse turbines, Sec. 35. 
(3) Both ends of reaction turbines. The water in a centrifugal 
water-packed gland must, when used in close proximity to 
steam which is above atmospheric pressure, be circulated and 
cooled to keep it from vaporizing. For a gland of this type 
that is used on the exhaust end of a turbine which is operating 
condensing, the water does not need to be circulated. 

116. The Gland Sealing Water For A Centrifugal Water- 
packed Gland Must Not Contain Any Sediment Or Scale- 
forming Salts. This is because if the water does contain such 
substances, the centrifugal action and the heat will cause the 
solids to be deposited in the gland in the form of scale. The 
scale will clog the gland and frequent disassembling and 
cleaning will be required. If scale is formed within the gland 
chamber and allowed to accumulate, the runner will eventu- 
ally rub and cause excessive vibration and leakage; or in 
extreme cases, the runner may be broken. 

117. The Arrangement Of The Gland -water Piping (Fig. 117) 
will depend upon the available supply of pure soft water. 



However, the general scheme which is usually employed, con- 
sists of a tank or reservoir, R, located at a sufficient height 
above the glands so that the proper water pressure in the 
glands will be provided by gravity. One such arrangement 
is shown in Fig. 117. Where the only available supply of 
pure water is that for boiler feeding, and the condensed steam 
is pumped directly back to the boiler, the gland-water reservoir 
may be supplied from the delivery of the condensate pump. 
In such cases, the gland-water reservoir should be of sufficient 


.Circulating Wafer 
From Condenser 

: i^^^^^^^^^^^^#^^^^^^. 


Fig. 117. — Piping arrangement for centrifugal water-packed glands. 

capacity so that the water which is delivered to it will have 
ample time to cool before it enters the glands. Where the 
water must flow through the gland (Sec. 115), the discharge 
may be piped to a feed-water tank or to the hot-well. 

118. Carbon-packed Glands (Fig. 118) may be used for 
packing the steam-end and the exhaust-end of turbines of all 
types, and also for packing the diaphragms of pressure-staged 
or of velocity-and-pressure-staged turbines. Carbon-packed 
glands which are used in the steam . and exhaust ends of 
condensing turbines are generally provided with a steam seal 
(Sec. 112). The steam which leaks through the glands and 
condenses must be drained away. Steam-seal piping, drain- 



'Pressure Gage ^1 
\ ', 'Gage Connect Ion 

-To Sewer- 
1-Carbon-Packeol G I an d s 

.Packing Springs 
^ And Holders 

-Stop Pin 

tl-Sec+ton Th.rou.g*h Carbon PacKmg Ring^ 

Fig. 118. — Carbon-packed glands in head end, exhaust end, and diaphragms of a 
pressure-and-velocity staged impulse turbine. The steam-seal piping is also shown. 
General Electric Cb.) 



age piping, operation, and repair of carbon-packed glands are 
treated in the following sections of this division. 

119. The Construction Of Carbon-packed Glands varies 
according to the conditions under which they are to be used, 
and also according to the manufacturer. Carbon-packed 
glands (Fig. 119) consists of one or more carbon rings which are 

contained in a chamber, C. The 
carbon rings encircle and fit 
closely to the shaft, S. They 
are made, usually, in three equal 
segments (Fig. 118-77) which are 
butt-jointed one to the other. 
These segments are held together 
either by a garter spring {B, 
Fig. 120) which completely en- 


Spiral Spring 
(Garter Spring) 

ITwo Chambers Each Con+alning| 
Two Packlnoj Rings 

^ . , Gland Chamber B 
Axial '■ P7: 


.. - -Turbine Casing - 
/' SfectmSecrl Space. 

Carbon '*• 


Jiingf No. 2 

K-Two Chambers Each Containing 
One Packinoj Rinqj 

Garter {~ 
Spring^ _i=ij 'Ji^* 


Connection'' \R> 

I- Transverse E-LongitudinaJ 

Section Section 

Fig. 119. — Showing various arrangements Fig. 120.- 
of carbon rings in carbon-packed glands. 

-Carbon-ring glands of the Terry 

circles the ring, or by three flat tangential springs (Fig. 118-77) 
which bear against the inner periphery of the chamber. The 
chamber is provided with one or more lugs C, Fig. 120, or 
straps which engage with a lug or keyway that is carried by 
the ring, thus preventing the ring from rotating with the 
shaft. Carbon-packed glands which are used in the dia- 
phragm of a pressure -staged turbine generally consist of only 
one ring. But the head- and exhaust-end glands may com- 
prise any one of various arrangements, (Fig. 119) such as two 


chambers containing one ring each, two chambers containing 
two rings each, three chambers, containing one ring each (Fig. 
120) etc. 

120. The Steam-seal Piping Of Carbon-packed Glands 
may be arranged as indicated in Fig. 118. The live-steam 
admission. A, is taken from the boiler side of the main throttle 
valve. The pop valve, C, is set to blow at a pressure of about 
10 lb. per sq. in. gage. If the turbine is to be operated non- 
condensing, the steam seal is not required, consequently the 
globe valve, F, may be closed. If the turbine is to be operated 
condensing, and the vacuum is established before the turbine 
is started, then both the head- and exhaust-end glands should 
be steam sealed. To effect this seal, valves F, D and E are 
opened so that the gages at M and N read about 3 lb. per sq. 
in., or so that a slight steam cloud issues from both packing- 
box-drain pipes, Di and D-z. Then when the turbine is brought 
up to speed, valve D may be closed. The packing-box drains, 
P and P, should be piped to a region wherein the pressure will 
never be above that of the atmosphere. Although there are 
other arrangements of steam-seal piping, they will not be 
treated in this book. 

Note. — The Steam Leakage At The Drains Of Steam-sealed 
Glands Should Preferably Be Visible From The Turbine Room 
as suggested at P, Fig. 1 18. Such an arrangement will enable the atten- 
dant to readily observe the amount of steam which is issuing from the 
glands. It is desirable that there be a slight leakage of steam (just a 
trace of visible water vapor) from carbon-packed glands. This provides 
a sort of lubrication for the carbon rings. Also, unless some steam is 
leaking from the exhaust-end gland of a condensing turbine, air is prob- 
ably leaking into the turbine. If the steam leakage from the exhaust-end 
gland is excessive when the gland-pressure gage reads about 3 lb. per 
sq. in., the carbon rings should be refitted. If an excessive amount of 
steam leaks from the head-end gland, these rings should be refitted. 
About the only way to determine whether or not diaphragm carbon 
packing (Fig. 118) needs refitting is, when the turbine casing is opened 
for inspection, to check the clearance (Sec. 121) with a thickness gage. 
Methods of refitting carbon packing rings are discussed in the following 

121. The Diametral Clearance Between A Carbon Ring 
And The Shaft should be about 0.002 in. per in. of shaft 


diameter when the shaft is cold. This will, due to shaft 
expansion, provide a total diametral clearance of approximately 
0.000,5 to 0.001 in. when the tm^bine heats up during operation. 
For high pressures and superheat, the diametral clearance 
should be about 0.003 in. per in. of the cold-shaft diameter. 
On small capacity turbines — up to about 100 kw. — the rings 
may be bored to approximately the cold-shaft diameter. 
Then, after two or three hours run, they will wear to normal 
size and an extremely accurate fit will result. However, this 
procedure is not advisable for large turbines because, if the 
rings pinch the shaft of a large turbine, serious heating and 
vibration may be caused. 

Note, — The Axial Clearance Of Carbon Packing Rings (Fig. 119) 
should be from about 0.003 to 0.006 in. That is, the width of the groove 
in the packing casing, as measured in an axial direction, should exceed 
the axial thickness of the carbon ring by this amount. If the clearance 
is too small, rust and sediment are Hkely to cause the ring to stick. If 
the clearance is too large, the steam pressure may not hold the ring tightly 
against the side of the groove, and steam will leak around the outside 
of the ring. 

122. A Mandrel Will Be Found Extremely Convenient In 
Fitting A Carbon Packing Ring. — The diameter of the mandrel 
should be the exact size to which the ring is to be fitted. The 
correct diameter may be determined by the amount of the 
required clearance as stated in Sec. 121. A piece of iron pipe 
can easily be turned to the proper diameter. The ring can 
then be easily and accurately fitted around this mandrel. 

123. In Refitting A Carbon Packing Ring which has worn too 
large, the inner diameter must be decreased. This may be 
done by filing off the joints (Fig. 121) and then reboring, as 
hereinafter explained so that the inner periphery of the ring 
will be a true circle of the proper diameter. When the rings 
are but slightly worn so that the diameter does not have to be 
decreased more than about 0.004 or 0.005 in. it is not necessary 
to rebore. The joint surfaces at the ends of the segments may 
be filed off and the ring assembled on the shaft. Then the 
shaft will wear the inner surface of the ring to a true circle. 
For methods of decreasing the inner diameter, see note below. 


Note. — In Filing Off The Surfaces Of The Joints Of A Carbon 
Packing Ring it is of paramount importance that the finished surfaces 
of each joint be true as shown in Fig. 122. To assist in filing a true sur- 
face, a wooden jig or pattern may be made and used as indicated in Fig. 

tegmenf Of 

Packinsr ' ^.Surface Square With 

**"^Sf., .-^=._ Jore And Used To 

Guide File 


fit between 

Carbon And 

Wooden Holder-'' 

Radius Same As 
That To Which 
Outside Of Packing 
Is Turned 

Fig. 121. — ^Wooden jig for holding car- 
bon packing rings for filing the joint 
surface. (E. H. Thompson in Power, 
Sept. 21, 1920.) 

I-P\an View 

31- Elevation 
Fig. 122. — Joint surfaces of each 
segment of a carbon packing ring 
must be made true. (The plane of 
surfaces A should be perpendicular 
to the plane of surface B. Surfaces 
A should also coincide with a radial 
line R.) 

121. If such a jig is not available, the three or four segments of a ring 
may be held in a vise as shown in Fig, 123 being careful to line up the 
joint surfaces of all of the segments. The relatively large area thus 
provided by the ends will assist materially in guiding the file. Be careful 

•Segmenti Of Rings 

Fig. 123. — Carbon packing rings fit- 
ted in vise preparatory to filing the 
joint surfaces. (E. H. Thompson in 
Power, Sept. 21, 1920.) 

.Carbon Packing Ring 


Fig. 124. — Hacksaw used to decrease 
the diameter of a carbon packing ring. 
(E. H. Thompson in Power, Sept. 21, 

not to screw the vise up too tightly as the rings are likely to be broken. 
If the vise jaws are rough, they may be lined with sandpaper. If the 
joint surfaces require a considerable amount of dressing down, the entire 
ring may be clamped on a board (Fig. 124). Then, with a hacksaw, 


cut through each joint, keeping the saw in a radial and vertical position 
so that the blade lines up with a diameter of the ring. 

Note. — The Carbon Ring Should, Usually, Be Bored Out To 
The Proper Diameter (Sec. 121) after the joint surfaces have been 
dressed down as explained above. The three segments of the ring are 
assembled and the boring done on a lathe. A large strong ring with its 
segments held together with the spiral spring or with a wire wound around 
its outer circumference may sometimes be held in the lathe chuck for 
reboring. But the best method is probably to make a wooden chuck 
by clamping a wooden block in the lathe chuck or in its faceplate, and 
then boring a cavity in the block, into which the ring will just fit. The 
bored surface of the ring should be made smooth by polishing it with No. 
00 sandpaper. Emery cloth should not be used on the packing rings 
because particles of emery will stick to the ring and then cut the shaft. 
If by accident the ring is bored out a little too large, the joints may be 
dressed down as explained above, and no reboring will be required. 


1. Why does the satisfactory operation of a steam turbine depend largely upon the 
condition of the shaft, bearings, and glands? 

2. How are the shafts of impulse turbines generally constructed? Of reaction 

3. What is meant by the critical speed of a turbine shaft? 

4. What is meant by a flexible turbine shaft? By a stiff shaft? Do most modern 
turbines have a flexible or a stiff shaft? 

5. What are the two principal types of steam-turbine bearings? 

6. Make a table showing the classification of steam-turbine bearings. 

7. Make a sketch of and name the principal parts of a -plain, flexible, steam-turbine 
main bearing. 

8. Make a sketch to explain the operation of a "flexible" bearing. 

9. In what kind of turbines are ball bearings sometimes used as main bearings? If a 
ball bearing becomes worn, what must be done? 

10. In general, what is the maximum temperature at which a main bearing should be 
operated? Name two means which are used to reduce the temperature of turbine 

11. What attention is necessary for the successful operation of a main bearing? 

12. Name five things which are likely to result from excessive wear of a bearing lining. 

13. Name two methods of repairing a turbine bearing. 

14. Explain with a sketch how a turbine bearing may be rebabbitted. 

15. What is the primary function of a steam turbine thrust bearing? 

16. Name four principal types of thrust bearings. 

17. Explain with a sketch the operation of the Kingsbury thrust bearing. 

18. What determines the axial clearance between the rotating and the stationary parts 
of a steam turbine? What is likely to happen if proper clearance between the moving 
and stationary parts is not maintained? 

19. What two types of mechanisms are generally employed for the axial adjustment of 
a turbine rotor? 

20. How is the correct axial adjustment of a tangential-flow turbine generally 

21. Explain how the axial adjustment of an axial-flow turbine rotor which is provided 
with adjusting screws is usually made. 

22. Explain how the axial adjustment of a turbine rotor is made with adjusting shims. 

23. Why must the thrust bearing itself have some clearance? 


24. How much clearance is usually allowed in a thrust bearing? 

25. What is a steam-turbine gland? 

26. What are the functions of a gland? 

27. Name four principal types of turbine glands. 

28. Make a sketch of a metallic-packed or stuffing-box gland. 

29. For what types of turbines and under what operating conditions are metallic- 
packed glands used? 

30. Explain how to repack a metallic-packed gland. Make a sketch of a skived joint. 

31. Explain how the wear in a metallic-packed gland may be taken up. 

32. Make a sketch of and explain the action of a metallic labyrinth gland. 

33. What are the two principal types of metallic labyrinth glands? 

34. For what purposes and in what kinds of turbines are double labyrinth glands used? 

35. Explain the operation of a steam seal. 

36. State the advantages and disadvantages of a double labyrinth gland. 

37. Describe the packing ring used in a single labyrinth gland. 

38. Explain how a single labyrinth gland may be refitted after it has become worn. 

39. What is a centrifugal water-packed gland? Explain its operation. 

40. Why cannot a centrifugal water-packed gland be used in close proximity to high- 
pressure steam? If a centrifugal water packed gland is to be used close to steam above 
atmospheric pressure, what means are employed to prevent the water in the gland from 

41. What must be the condition of the gland sealing water? 

42. Upon what will the arrangement of the gland- water piping depend? 

43. Where may carbon-packed glands be used? 

44. Make a sketch showing one method of steam-seal piping for carbon-packed glands. 

45. Why should the steam-seal drains be visible from the turbine room? 

46. What diametral clearance should be provided between a carbon packing ring and 
the shaft? What axial clearance should be allowed? 

47. Explain with sketches how to refit a carbon packing ring. 



124. A Steam-turbine Governor Or Speed Governor Must 
Be Used Whenever It Is Desired To Have A Steam Turbine 
Run At A Constant Speed While The Load Which It Is Driving 
Or Its Rate Of Doing External Work Or The Supply-steam 

btoam Supply--' 

Worm-.., Flyb^^l^o^^rnor 

Vctlve --■ 

"Spent" rCca^vf;W 


Fig. 125. 

-Governor used on De Laval vertical oil-purifier turbine which is of the 
impulse type. (De Laval Separator Co.) 

Pressure Varies, Fig. 125 (see Sec. 27 for definition of 
governor). If steam were constantly admitted at the same 
rate to a turbine while the resistance to the turning of its rotor 
(due to the external load) changed considerably, its speed would 
fluctuate excessively. A very great load might stop it. A 
sudden decrease in the load would allow the speed to increase 
to a dangerous value. Obviously, if the speed of the turbine 
is to be maintained constant and unless the admission of steam 
is controlled by hand, there must be some automatic means of 



proportioning the steam supply to the varying load on the 
turbine and the varying pressure of the steam supply. 

Note. — In Marine Service And In Driving Blowers, It Is Possi- 
ble To Operate A Turbine Without A Speed Governor. — In such 
service, the resistance (torque) which the propeller or blower offers to 
the rotation of the turbine increases with the speed. The work which the 
turbine does increases faster than the speed. The turbine will therefore 
find a certain constant speed at which any given steam supply will be 
sufficient for the work done. In most stationary services there is a 
possibility that the load may be suddenly removed entirely. Then, the 
only limit of the turbine rotor's peripheral speed would be that equiva- 
lent to the velocity of the steam jet, which is usually high enough (Sec. 
10) to burst the rotor due to the centrifugal force. 

Valve Open-- 

Fig. 126. — 'Diagram of direct throt- 
tling governor for a steam turbine. (The 
imaginary construction here shown is 
never used in practice.) 



Fig. 127. — Diagram of the same imagi- 
nary governor as Fig. 126 but in the 
closed position. 

125. How A Governor Keeps The Speed Of A Turbine 
Nearly Constant, in spite of considerable variations in load, 
may be understood by a study of Figs. 125 and 126. 

Explanation. — Figure 126 shows an imaginary turbine governor. The 
steam flows through the nozzle, //, and impinges on the buckets of rotor, 
A, causing it to rotate. The movement of the rotor shaft is reduced and 
transmitted through worm gear, B, shaft, C, and bevel gears, D, to the 
spindle of a fly-ball governor. The weights, E^ of the governor rise due 
to centrifugal force (see the author's Steam-engine Principles And 
Practice). The vertical movement of the weights is transmitted 

Sec. 126] 



through the drop-rod, F, to butterfly valve, G. If the speed of the rotor 

increases beyond a certain value, the weights will fly out so far that the 

valve will be entirely closed as in Fig. 127. Then the speed of the rotor 

will naturally decrease for lack of steam. The weights then fall and 

more steam is admitted as in Fig. 126. In this way, the governor being 

properly designed and adjusted, the turbine is prevented from running 

much faster or much slower than its rated speed. 

Note. — The Speed Regulation Of A Turbine is the ratio of the speed 

decrease from no load to full load to the full load speed. Or, expressed 

as an equation: 

{No-load speed) — {Full-load speed) 

(27) Speed regulation = „ „ , — (decimal) 

Full-load speed 

126. A Complete Goveming-mechanism For A Steam 
Turbine consists of several parts. There is always a centri- 

Overspeed .-Knife Edge Block 
•Weight \ .'K nifeEdge 


Ball Thrust 
: Bearing On End 
\ Of Governor Lever 
Governor Spindle 
^•Governor Spring 
""Governor Y/eigtit 

Fig. 128. — Governor of Moore steam turbine. (Instruction Card No. 2.) 

fugal device (Fig. 128) or rotating part commonly called the 
governor proper. This device usually consists of movable 
weights so mounted that they are acted on by centrifugal 
force and, in some designs, by inertia also. An exception 
to the general construction is the pneumatic governor of 
the Ridgway turbine shown in Fig. 157. This governor 
mechanism has a pressure blower directly connected to the 
shaft instead of the usual movable weights. Since the pres- 
sure developed by the blower varies with its speed, the blower 
pressure can be used to regulate the speed of the turbine 
(see Sec. 148 for a description of the operation of this type 
of governor). There is always also a valve or a number of 


valves (Fig. 129) in the steam passage leading to the turbine 
nozzles which valves are in some way controlled by the cen- 
trifugal force of the weights. Between these two essential 
elements (the governor proper and the valves) there is a 
connecting mechanism of some one of the many kinds which 

Steam Governor Valve 
Chest -.^ Seat Bushing 

Lever - 

Butterfly Valve 

Fig. 129. — Steam chest of Moore steam turbine (Instruction Card No. 2). The 
governor valve, F, is operated by the governor (Fig. 128) through lever, L, and valve 
stem, (S. Valve, F, is of the balanced type which has seats in a valve bushing, B. A 
steam-tight joint is made between the end of B and the steam-chest cover, C, with 
asbestos packing soaked in graphite and oil. Metallic valve-stem packing is used. A 
lantern gland is used from which the leakage along the stem can be piped to the atmos- 
phere. The relative position of the valve is fixed when the turbine is tested and the 
valve stem nuts, N , are pinned to the shaft. This adjustment should not be changed. 

are used by various manufacturers for transferring the 
governor-weight motion or blower pressure to the admission 

Note. — Steam-pressure And Oil-pressure Governor Systems are 
employed, as is explained later, on large turbines through pilot or relay 
valves. These pilot or relay valves multiply the force derived from 
the governor proper and are necessary because, in large turbines, the 
force required to move the governor valve is so great that it is impracti- 
cal to operate the valve directly by the governor proper. Governors 
which employ such systems are called relay governors. Many different 
mechanisms are also in use which obviate part of the losses of available 
energy which result from throttling the steam at light loads through 
valves which are "cracked" or nearly closed. One of these mechanisms 
which admits the steam to the turbine in "puffs" is described in 
Sees. 136 to 138; another which admits the steam through a multi- 
ported valve, in Sees. 144 and 145. The term governor is used in the 
following table to indicate a complete governing-mechanism and not 
merely the governor proper. 

Sec. 127] 



127. Table Showing The Various Ways In Which Turbine 
Speed Governors May Be Classified and the various sub- 
classes under each classification (see preceding note). (These 
classes will all be explained and illustrated in the following 


Classified with 
respect to 


Class or 



Actuating force. 



Fig. 136 


Centrifugal and inertia. 

Fig. 156 


Air pressure. 

Fig. 157 


Method of valve 



Fig. 130 



Mechanical indirect. 
Indirect or relay: 

Fig. 159 


(a) Steam relay. 

Fig. 143 


(6) Oil relay. 

Fig. 142 


Valve arrangement 



Fig. 130 

or steam control. 


Varying nozzle area 
(multiple valve). 

Fig. 152 



Fig. 147 



Fig. 145 

128. A Direct Centrifugal ThrottUng Governor (classes 1, 4 
and 8, Table 127) operates as explained in Figs. 126 and 127. 
Governors of this type are widely used on small turbines. 
They are very simple as compared with some of the other 
types and are, on the whole, very reliable. The throttling 
action of the control valves of governors of this type decreases 
the efficiency of the turbine somewhat at light loads; it is to 
avoid this loss in efficiency that other methods of steam-flow 
control are employed in governing. Some commercial 
governors of this type will be explained in the following sections. 

129. The Main Governor Mechanism Of The Sturtevant 
Turbine shown in Fig. 130 represents one commercial applica- 
tion of a governor of the direct centrifugal throttling type. 
The spindle of the governor is horizontal and the movement 
of the centrifugal weights is opposed by a single heavy spring. 


Explanation. — As the speed increases, the centrifugal weight arms 
(A, Fig. 130) "fly out," forcing the governor spindle, B, against the 
ball-bearing socket, C, located in the head of the bell-crank lever, D. 
This motion is in turn imparted, through the eye-bolt, M, to the valve, Z, 
which, in rising, closes the steam ports. When the speed decreases, the 
action is as follows: The weight arms, A, are drawn in by the centripetal 
force of the main governor spring, E. The external spring, P, pulls 
down on the end of the bell-crank, D, causing the ball-bearing socket, (7, 
to follow the inward movement of the governor spindle, B, thereby 
lowering and opening the valve, Z. In other words, the governor closes 
the valve and the external spring, P, opens it. The tension of this spring, 

I k- V Centr ifugal- Weight Arms 

: Bel /-Crank lever J) 

"^^■^=2^ .-Governor Spindle^ 
— '^^ Bat IBeanng Socket^ 

Lock Nut ^ 
Thrust Bearing^ 
Adjusting NutW 

•ustlng Nut\, 
■External Spring P 

y^-\/alve r 

i^ Spindle^ ^ * 

Strip Or Key ^. 
In Position-': 
-Steam In let 

I-Genercul Assembly I- Ball Bearing Socket 
Fig. 130. — Main governor mechanism of Type-6 Sturtevant turbine. 

P, is varied by the adjusting nut, L. The dust shield, A^^, prevents dust 
or grit from working in around the spindle, 0, which would increase its 

130. To Adjust The Sturtevant Governor Valve (Z, Fig. 
130), proceed as follows: Insert a strip or key ^{q in. thick 
between the governor spindle, B, and the ball-bearing socket, 
C, 'as shown at X. This is done with the throttle valve 
closed. After opening the throttle valve, the block, X, being 
in position as shown, the steam gage on the turbine steam chest 
should then show a slight pressure, say 10 or 15 lb. per sq. in. 
If there is a higher pressure than this, the valve, Z, should be 
raised by adjusting the eye-bolt M; in other words, remove the 


bell-crank, D, loosen the dust shield N (which also acts as a 
lock-nut) and screw the eye-bolt, M, on the valve spindle, 0. 
If no pressure shows on the turbine-steam-chest gage or the 
pressure is too low, it can be increased by lowering this 
valve. This setting will give the maximum opening for full 
load and will, at the same time, prevent overspeeding at 
light or no loads. The valve adjustment can also be made 
at thrust-bearing body, F, by firstr loosening the locknut, G. 

Note. — The Thickness Of The Stock Used Between The Gover- 
nor Spindle, B, And The Ball-bearing Socket, C, as shown at X, 
varies for different types or turbines. On turbines equipped with a ball- 
bearing step {F, Fig. 130) on the governor pin end use 3'^2-in. stock for 
types A-6 and B-6, and 3^ g-in. stock for types C-6, D-6 and E-6 turbines. 
On turbines which {ire not equipped with a ball-bearing step on governor 
pin end, use ^g-in. stock on all types. If the governor seems to "jump" 
or remains unsteady, and thus interferes with the operation of the tur- 
bine, this can be eliminated by adjusting the lower valve disc. It may 
be necessary to make several trials in order to determine the correct 
location of the valve disc. In making this adjustment the disc should 
not be moved more than }^ of a turn at one time and, of course, should 
be securely locked after each adjustment. 

Note. — Adjustment For Change In Speed Of The Sturtevant 
Governor (Fig. 130) may be made by adjusting the nut, H, in the end 
of the governor. The speed will be increased by screwing in the nut 
and lowered by backing it out. 

131. A Direct Centrifugal Throttling Govemor Which Is 
Provided With An Auxiliary Vacuum-breaker Attachment is 
shown in Fig. 131. When a turbine is operated condensing, it 
may be necessary to break the vacuum in order to prevent 
racing when the load is removed suddenly. 

Explanation. — If the nut, D, which is deflected by the movement of 
the governor, travels outward more than about 3^^ in., it engages the end, 
/, of the hollow valve stem, T. The movement of T admits air to the 
turbine exhaust passages through ports O and P. 

132. Other Direct Throttling Governors are shown in Figs. 
132, 133, 134, and 135. That in Fig. 135 is almost identical 
with the leaf-spring governors used for small steam engines. 
(See the author's Steam-engine Principles And Practice.) 
The following instructions for care and adjustment of these 


governors may be applied to almost any small governor of the 
direct throttling type. 

Note. — Speed Adjustments Op Direct Throttling Centrifugal 
Governors, to provide a speed 2 or 3 per cent, greater or less than the 
existing speed may always be made on governors of the coil-spring type 
by screwing up or slacking off on the main-spring tension. Very slight 
changes in speed may also be secured by varying the external spring 
tension. In changing the tension on either of these springs, care should 

<r---Sfecrn7 Suppflf 

Fig. 131. — De Laval governor equipped with vacuum breaker. 

be taken to prevent the springs from becoming "coil-bound" — entirely 
closed — when in service. For any material change in speed, in governors 
not provided with regular speed-changing handwheels, it is best to con- 
sult the manufacturers who will usually supply new weights or springs 
for the new speed. After any adjustment, the governor should be 
examined, moved by hand or watched to make sure that it shuts off at 
no load and moves freely in all positions. 

Troubles Of Direct Throttling Governors are generally hunting 
or racing due to sticking of the mechanism or faulty adjustment. Lost 
motion will also cause hunting. Lost motion may be taken up in the 
valve stem (Fig. 132) of some governors. The lost motion may usually 
be detected by moving the various parts and observing the fit. A 

Sec. 132] 



certain amount of lost motion in the stationary position is sometimes 
recommended by the manufacturer. This lost motion must not be so 
great as to prevent the governor shutting off, A sticking valve stem may 

'Oovemor Valve Box 
''Governor- Valve Bonnet 

Fig. 132. — Governor valve of Terry turbine. 

usually be detected by pushing the valve in and noting if it springs back. 
If the valve does not shut off at no load and thereby allows the turbine 
to race, it probably leaks or its stem is too short. The effective length 
of the stem can be increased by means of adjusting nuts. The cause of 

,. Oil And Grease Cup 


.' Oovernor- 
\ \ Weight Knife Ecfge 
\ ^Governor Slide 
Oovernor Ac/Justing Nut 


Fig. 133, — Governor of Terry steam turbine. (The shaft, A, supports the governor 
disc, B, by means of a taper shank which is keyed in position by taper pin, L. The 
governor weights, C, are supported on knife edges, Z). The weights move the governor 
sUde, H, outward by means of the yoke, G, against the tension of spring F. The move- 
ment of the slide is communicated to lever, P, by means of slide end, M, which revolves 
against ball, iV. Oil is fed by Q to the ball thrust. The governor is housed in S. The 
main speed adjustment is by nut Ri) 

leaks should be investigated. If due to rust, the valve can be cleaned 
to insure a better seat. Conical-seated valves may be refinished on a 
lathe and "ground in" by an experienced machinist. Corrosion of the 
valve is prevented by keeping the turbine well drained when it is idle. 


Note. — The Following Possible Causes Of Governor Hunting 
are given by the Westinghouse Electric & Mfg. Co. for the direct throt- 
tling governors on their mechanical-drive turbines. (1) Too great a travel 

.'^ Sfandarzf 0/7 Cup 

;Ball Thrust Bearing. 
' ,&overnor5lide. 
: ,'Wheel Shaft. 
/ • Slotin 

', HolhvtShaft 

fr. ■ Governor Slide End Nut. 
■-Governor Lever, 
'''Jf"Standard Pipe Plug; Remove For 
Taking Speed With Tachometer. 

Fig. 134. — Ball thrust bearing in governing mechanism of some Terry turbines. 

of governor poppet valve. (2) Sticking of governor poppet valve on 
guide. (3) Sticking of governor spindle. (4) Bent valve stem. (5) 
Broken governor weight knife edges. (6)- Distorted or bent governor 
linkage. (7) Weakening of governor springs. 


>5team Chest ■ 
Fig. 135. — Governor of Steam Motors Company turbine. 

133. The Emergency — Or Overspeed — Governor Mechan- 
ism Of The Sturtevant Turbine (Figs. 136 and 137) operates 
only in case of failure of the regular speed governor. When 

Sec. 134] 



the turbine is running properly, the speed is controlled or 
governed by the speed governor; that is, the turbine is said to 
be ''running on the governor." But should the governor lose 
control of the turbine (permitting it to run too fast) there is 
danger of accident unless some safety device, which will act 
automatically, is provided to ''shut down" the turbine. To 

'No2zle Valves 


Fig. 136. — Emergency- and main-governor-mechanism assembly of Type-6 Sturtevant 


avoid this danger, the emergency governor is provided. See 
explanations under Figs. 137 and 138. 
l/ 134. To Adjust The Emergency Govemor (Figs. 136 and 137) 
screw in or out on the adjusting plug, which is located, opposite 
the point where the piston. A, protrudes. Screwing this plug 
alters the relation of the piston's center of gravity to the center 
of rotation. Consequently, the closer the center of this plug 
is to the center of the shaft, the higher will be the speed at 
which the emergency governor will operate, and vice versa. 
Do not make the mistake of adjusting the stop bushing which 
holds the piston spring in position, for this will change the 


distance which the piston extends when it flies out. The 
clearance between the tripper, B, and the rotating element 

r Adjusting Plug 


:" ji^^iearance 
( 'Tisfon^A 
''Stop Bushing 



''Emergency Valve =E 
I-Sectiona! View H-Side Elevation 

Fig. 137. — Emergency-governor mechanism of Type-6 Sturtevant turbine. When 
overspeeding, the piston, A, "shoots out" and strikes the tripper, B. B then causes 
the bell crank, C, to release the valve lever, D, which is directly connected to the emer- 
gency valve E, thereby causing E, to close. E is kept open by the valve lever D being 
held up by the bell crank C against the action of a strong valve spring F. When D is 
released, the strong spring comes into action, causing the rapid closing of the valve. 

should not be more than Jfg in. If the emergency governor 
trips, it cannot be reset until the speed of the turbine has 

■ Governor Cover 

Compression Spring P=Overspeect 

' Trip -Lever Siiaft 
I-Section A-A 

;' ''Governor 

'^Turbine Shaft 
Trip Lever 

I- End Sectional View 

Fig. 138. — Overspeed governor, Moore steam turbine. (Instruction Card, No. 2.) 
This overspeed governor consists of a small pin, P, which is held in place by a compres- 
sion spring. At a certain predetermined speed, for which the governor is set, this pin 
is thrown out and trips a latch, L, operating a butterfly valve, F, which cuts off the 
supply of steam to the turbine. See also Fig. 128 for another view of this emergency 

decreased to about one half of its running speed. This action 
is caused by the pin being unstable and moving to its limit 
when once started. The emergency governor should be 

Sec. 134] 



adjusted to trip at about 10 per cent, above the normal 
running speed. The emergency governor should be tested 


Depressions To Hold Spring 

\ Slot; .-Motion Limiting Stud 



Emergency- Vaive 
Operating S/?a ft'. 

Vaive Operating 

Governor ^ 
Vise -'' 

Fig. 139. — Ring-type emergency 
governor used on the smaller Terry 

Fig. 140. — Pivoted-lever type of emer- 
gency governor on Terry turbines. 

Turbine Shaft 
Governor Weights 
n-Sidc Elevation 

Fig. 141. — Emergency governor oi Steam Motors Com-panytxahva.^. (Steam Motors 
Company, Springfield Mass.) The emergency governor is a device for shutting down 
the machine in case of a "runaway." It is not a speed-regulating governor. The 
governor weights, TF, are so adjusted that when the turbine shaft attains a speed 10 per 
cent above the maximum operating speed they will "fly out." They then strike trigger, 
T. This trigger releases lever L, which gives a hammer blow to rod R, releasing the 
other tripping mechanism on the valve bonnet. The emergency valve will then be 
closed by spring S. To reset this emergency trip, lift M, set N , in position and replace 
the catch T. 

periodically^ by holding the governor rod against the force of 
the centrifugal weights, until a 10-per cent, overspeed is 


obtained as shown by a voltmeter or reliable tachometer. It 
is important that the overspeed governor mechanism be 
always ready for an emergency. Nearly all emergency 
governors may be adjusted to trip at a lower speed by 
moving the weight further from the center of rotation. 

-Oil Cylinder 

Oil Pump- 

's. '•■0/IUncfer 
:■ 3£To40Lb. 
I Pressure 
Bearing Case 

'"eovernor- Supply 
Steam Valve 
^'' Steam Chest 


Fig. 142. — Oil-relay governor and steam chest of Moore steam turbine (Moore Steam 
Turbine Corporation, Wellsville, New York; Instruction Card No. 3). A governor, G, 
is used to actuate the oil-relay control. An increase of speed causes the weights, W , to 
move outward. This moves lever L upward, moving oil-relay valve, Y, which admits 
oil below piston, P. This causes the governor steam-valve, <S, to close. Movement of 
8 moves compensating lever, C, which brings Y back to its neutral position. This stops 
the flow of oil and prevents over travel of the steam valve. The governor steam valve, 
jS, is provided with a spring, M, at the lower end of its valve stem. The purpose of this 
spring is to automatically close the valve and shut down the turbine in case the oil pres- 
sure fails. The overspeed governor, O, is carried on the governor shaft above the worm 
wheel, X, which drives the governor; a weight is held in place by a compression spring 
until a predetermined speed, for which the overspeed governor has been set, is reached. 
Then the overspeed governor is thrown outward and strikes a lever, H, which trips a 
latch, allowing auxiliary valve, vl , to be forced upward by spring B. This admits full 
oil pressure under piston P and exhausts oil from above the piston, closing the governor 
steam valve, S. 

Note. — In Maintaining The Emergency Governor (Fig. 137) the 
following should be observed. The piston, A, should "shoot" out at a 
speed about 10 per cent, greater than the rated speed of the turbine. 
This piston should occasionally be tested for free movement. To make 
this test, push a wire through the hole in the center of the adjusting plug; 
it should be possible to thus push the piston out approximately )^ in. 
It is very important to have this piston working freely, and a little oil 
applied occasionally — say once a month — will assure this free movement. 

Sec. 135] 



Note. — Other Makes Of Emergency Governors are shown in Figs. 
138, 139, 140 and 141. Their actions and functions are similar to those 
already described. In general, the emergency governor should be 
entirely independent of the speed governor. 

Fig. 143. — Diagram showing operation of the older-type Parsons turbine governor. 

135. An Oil -relay Throttling Governor (Fig. 142), accom- 
plishes the same result as does the direct throttling governor 
but does not depend on the centrifugal force of the weights to 
operate the main governor valve. Instead, the centrifugal 
force of the weights operates a small valve which admits oil 
above or below a piston the rod of which controls the main valve. 


136. Centrifugal Steam-relay Intermittent Or *'Blast»» 
Governors are used on a large number of Westinghouse and 
foreign Parsons turbines. Oil-relay governors (Sec. 138) 
are superseding this type. The principle of operation of the 
governor may be understood by examination of Fig. 143. 
Its action is, briefly, to admit steam to the turbine nozzles 
in ''puffs," the length of the ''puff" depending on the load. 
The "puffs" occur at regular intervals and so frequently 





Fig. 144.- 

■Absolufe Zero 
' " - - /4 fmospheric =0 
-Graphs showing the effect of an intermittent governor on the instantaneous 
steam pressure in turbine live-steam parts. 

that there is no uneven effect on the speed of the turbine. The 
principal object of this action is to have the valve either 
entirely closed or wide open most of the time, so that there will 
be little throttling. Another advantage is that, since the valve 
is constantly moving, the possibihty of its "sticking" is mini- 
mized. With the advent of the larger turbines this "puff" 
system of admitting steam was found to cause, at times, 
objectionable vibration in the main steam lines of the power 
house. About 1909 the steam relay began to be abandoned for 
the oil-pressure-relay system. 

Explanation. — The turbine shaft (Fig. 143) carries a worm, W. The 
shaft of the worm wheel which engages W carries an eccentric, E, and a 
bevel gear, fi, which drives the spindle of the centrifugal governor, G. 
There is a system of levers connected to the eccentric rod, R, through 

Sec. 137] 



which it gives a reciprocating motion to the plunger of the relay valve, V. 
The live steam is admitted at N, flows through the space, Q, around the 
piston rod, C, and lifts the piston, P, which controls the governor valve, 
T. This allows steam to flow through T to the turbine as long as the 
valve, V, is closed. But when V is open, the steam escapes at M (into 
the engine room) faster than it enters at Q; thereby the piston is forced 
down by the spring, A, which presses behind it. One of the levers, L, is 
pivoted on the sleeve, S, of the governor so that when the governor lifts, 
V moves between higher limits and allows steam to escape at M for a 
longer period. In this way, the valve, T, is made to remain closed longer 
when the speed of the turbine is higher. The effect of this action on the 
steam pressure is shown in Fig. 144. 

CpnnecHng Fiocf To, Bypass yalre 

Governor _ , r , 

Oil-Pelay Synchroniiing 

'Cylinder ,^;/ Lever----, 

^^Synchronizing Handwheel 

Fig. 145. — Throttling and bypass governor used on Allis-Chalmers reaction turbines. 
See Fig. 146 for an enlarged view of the oil-relay valve. 

137. An Allis-Chalmers Oil -relay Throttling And Bypass 
Governor which is used by that company on 5,000 to 15,000 
kw. turbo-generators is shown diagrammatically in Fig. 145. 
Its action is similar to that already described for oil-relay 
governors in Sec. 135 except for the bypass and synchronizing 

Oil Outlets 
To Governor- 
: Piston "'^ 


Explanation. — As the turbine speed increases, weights W fly outward 
and raise the vertical rod, R, which is attached to the floating lever, D. 
This lever D, being supported at pivot C, pushes down on the stem, T, of 
the oil-relay valve (Fig. 146) thus opening its ports so that oil pressure 
is admitted abovepiston P. This closes the governor valve, U. But as 
U moves down, it moves compensating levers, E and G, and thereby 
moves upward F and synchronizing lever, A, which is pivoted at the fixed 

point K. Lever A is attached to D 

Relay-Valve ^^ P^^°* ^' "^^^^ movement in turn 

Rod raises T and closes the relay-valve 


When the turbine speed falls, due 
to an increased load, the above procr 
esses are reversed and the valve U 
is lifted from its seat. Its motion is 
communicated through connecting- 
rod, <S, to the sliding collar, N. At 
a certain position of JJ , the sliding 
collar strikes the fixed collar, Af, 
and the bypass valve, Y, is lifted. 
This admits live steam to an inter- 
mediate stage of the turbine. Thus 
the bypass valve remains entirely 
closed at light loads and opens for 
heavy loads. The end K of the 
short synchronizing lever, A, which 
is pivoted at C to the floating lever, 
Z), may be raised by screwing up 
on the handwheel, U. This changes 
the position of the relay valve with 
respect to the main governor valve 
and so changes the speed of the 
turbine. A 5-per cent, regulation 
above or below normal speed may thus be obtained. 

Note. — Bypassing Is Employed In Many Large Modern Multi- 
stage Turbines as a means of carrying overloads. The steam which 
is bypassed to a later stage of the turbine is not used with as high an 
efficiency as that which flows through all of the blading. There is there- 
fore, at overloads, a loss in efficiency due to bypassing but this loss is 
offset by the increased ability of the turbine to carry peak loads. Thus, 
for example, a turbine which operates at its best economy at 5,000 kw. can 
readily, by bypassing, be made to carry 7,000 kw. But when carrying 
7,000 kw., its economy is not as good as when it is carrying 5,000 kw. 

Fig. 146. — Enlarged view of the Alliz- 
Chalmers oil-relay valve shown in Fig. 

138. The Westinghouse Type Of Centrifugal Oil-relay 
Intermittent Governor is shown in Fig. 147 and the valves 


which it actuates in Fig. 148. (Based on Westinghouse 
Electric & Mfg. Go's. Instruction Book No. 5,171.) In 
general, the functions of this governor (the details of operation 
are given below) are: (!) To provide a throttle valve, (Fig. 148), 
which will be controlled by the governor proper for maintaining 
a constant turbine speed from no load up to about full load. 
This is effected by means of an oil-relay system, similar to that 
already explained in Fig. 142. (2) To provide an overload 
bypass valve, P (Fig. 148) , which opens at about full load and 
admits additional steam to a later stage of the turbine to carry 
overloads as explained in the preceding section. (3) To 
provide a continuous reciprocating motion of the throttle valve, 0, 
and the bypass valve, P, when the latter is open and of the operating 
linkage, whereby: (a) Sticking due to starting friction is avoided, 
(b) Energy loss due to throttling of the steam at very light loads 
is avoided. Z (Fig. 147) is the governor proper whereby the 
steam flow to the turbine blading is controlled by governor 
valves, and P (Fig. 148), which are, as will be explained, 
actuated by oil under pressure as regulated by the relay-valve 
system, FE. 

Explanation. — The worm, W (Fig. 147), mounted on the turbine 
shaft, drives a worm wheel which is mounted on the governor spindle. 
The governor proper is thus rotated. The cam, X, is driven by a gear 
on the governor spindle. This cam gives a rocking motion to the short 
lever, N, which is pivoted at q on the governor lever. In this way a 
short regular reciprocating motion, for reasons previously indicated, is 
transmitted through the linkage, MYSJ, to the oil-relay valve, E. See 
Fig. 149 for an enlarged view of this valve. As the governor raises it 
rotates lever / around its pivot e and hence lowers the rocking-lever pivot 
q. This causes the cam, X, to move the relay valve, E, between lower 

This oil-relay valve acts similarly to a piston slide valve for a steam 
engine. When raised it admits oil, from the pressure chamber, H (Fig. 
149), to the under side of the operating piston, F, simultaneously allowing 
oil to flow from the upper side of the piston to the exhaust passage, /. 
When E is lowered, its action is the reverse and the oil is admitted above 
and exhausted below the operating piston. The floating lever, G, to 
which the stems or rods of both the oil-relay valve and the operating 
piston are attached, operates to stop the oil flow as soon as the operating 
piston has moved a short distance. This lever is arranged in this way 
so that the operating piston will not move its entire stroke for only a 
small movement of the oil-relay valve. It is desired that the movement 


of the piston be proportional to (but much greater than) the movement 
of the relay valve. 

The operating piston, F (Fig. 148), controls (by the movement which 
it derives from the oil pressure through the pilot valve as explained 
above) the two valves — a primary or governor valve, O, and a secondary 
or bypass valve, P. The levers, m and n, which connect the operating 
piston to the valve are similar except that the lever, n, is provided at R 
with an adjustable amount of lost motion so that valve, P, will not lift 
until valve, 0, is open (see Sec. 141 for adjustment). The valves are 

fiufomaf/'c Safety; 

Main Goyernor Spring. 

Coyernor Ball- - -V 

Piyof Of nockingl 
Leyer On Ooyernor •; 
Leyer- . 

* ^Operating 



Speecf Changer 
Or Synchronizing. 



^-Rocking Car: ^'Oears 

- - -Limit Switch 

Fig. 147. — Operating gear and governor proper of a Westinghouse intermittent governor. 
The valves which this governor controls are shown in Fig. 148. 

provided with main springs, C, which close them if the oil pressure fails — 
as for instance when the turbine is stopped. When shutting these valves, 
the governor tends to raise the operating piston and would, when the 
governor is not revolving, strain the linkage if it were not for the weak 
spring, S (Fig. 147). This spring is inserted in the connecting link so as 
to permit closing the governor valve without straining the linkage. 

Note. — An Automatic Stop Valve, Q (Fig. 149) is provided to shut 
the governor valve in case of failure of the governor linkage. This valve 
consists of a piston, L, held to the top of a small cyHnder by the steam 
pressure on its unequal upper and lower faces. Live steam is admitted 
at U above the piston but leaks past and establishes a pressure in the 
lower part of the cylinder as long as the opening, V, is closed. The 


opening, V, is connected to the emergency governor (Fig. 150). When 
the emergency governor is tripped, it releases, through a pipe, the pres- 
sure in V' and the live steam at U then forces the piston, L, to the bottom 
of its cylinder against its spring. The movement of L throws a piston 
valve, T, which operates just as does valve, E, to close the governor 

139. To Check The Adjustment Of The Westinghouse 
Centrifugal Governor (Z, Fig. 147) first adjust the speed 
changer spring, d, so that it will have practically no tension 
when the governor balls or weights are in their innermost 
position. The main governor spring (which is held by nut, a) 
should now be adjusted so that the turbine will run at 5 per 
cent, below normal speed at no load. Then tighten d until 
the speed of the turbine is normal. There should now be the 
proper amount of speed regulation — about 1 per cent, between 
no load and full load. If there is not, then, for less speed 
regulation, adjust the nut, a, so as to render more coils of the 
main spring effective; for more speed regulation, so adjust that 
fewer of the spring coils are effective. 

Note. — Speed Adjustments While The Turbine Is Running are 
made by means of the spring, d. The wheel which tightens or loosens 
this spring may be so arranged as to be turned by a motor, which is con- 
trolled from the switchboard, so that the turbine may be synchronized 
with another one for parallel operation. 

140. The Oil-relay Control Adjustment Of The Westing- 
house Oil-relay Intermittent Governor (Figs. 147 and 149) 
should be made after the governor proper has been adjusted, 
as described in the preceding section. The method is as 
follows: With the oil -relay control connected and the oil 
pressure established, permit the turbine to turn slowly under 
steam so as to make lever, N, oscillate. The governor balls 
or weights should be in their innermost positions. Manipulate 
the oil-relay valve, E, by holding down on the pivot, J, to 
bring operating piston, F, into mid-position. Then adjust 
link, r (Fig. 149), so that when oil-relay valve piston, E, is in 
mid-position and will not admit oil either above or below the 
operating piston, F, the lever, G, will be horizontal. Then 
release J so that the spring link, S, is at its full operating 
length (not compressed) and the piston F, will move to its 


extreme bottom position. Now adjust link, M, until the 
piston, F, has a slight movement. Finally, lengthen M by 
giving it one and one-half turns. 

141. The Setting Of The Primary And Secondary Inter- 
mittent-governor Valves Of The Westinghouse Turbine 
(Fig. 148) may be checked as follows: The amount of travel 
of the valves from their extreme positions to their mid-posi- 
tions, when the levers m, and n, are horizontal, should be noted 


Valve - ~ 

Fig. 148. — Westinghouse operating cylinder, primary and secondary valves controlled 
by the governor of Fig. 147. 

at the time the turbine is delivered as complete by the erector. 
These travels should be afterwards maintained. With the 
primary valve, 0, just leaving its seat, the piston, F, should be 
y^ in. from the end of its stroke. This may be adjusted by 
inserting liners at point, Ifi. When the piston, F, is in its 
extreme upper position there should be from 3^:32 to 3^f e iii- 
clearance underneath link block, Z. This may be adjusted 
by inserting liners at point, g. The adjusting screw, R, 
should be so adjusted that the secondary valve, P, will open 
at the moment the primary valve, 0, reaches its maximum 
port opening, as shown by the pressure in the space, /. 

Sec. 141] 



Fig. 149. — Enlarged view of the Westinghouse relay valve of Fig. 147. 


142. The Automatic Stop Adjustment Of The Westinghouse 
Intermittent-governor Turbine may be checked as follows: 
With the automatic stop piston, L (Fig. 149), at the upper end 
of its stroke, the enlarged parts, B and C, of the safety stop 
plunger, T, should be central over the ports, A and A\, With 
the automatic stop piston, L, at its lowest position, the 
enlarged parts of the safety stop plunger, T, should be central 
over the ports, D and Bi. 

143. A Westinghouse Safety Stop Or Emergency Governor 
(Fig. 150) releases steam pressure in a pipe when it trips and 

Cove r nor Booly 
Screwed On 
The End Of 
Turbine 5haft\ 

Fig. 150. — Automatic emergency governor or safety stop which is used on some West- 
inghouse turbines in connection with the throttle valve of Fig. 151 and the valve, T , of 
Fig. 147. 

this drop in pressure operates one or more automatic valves 
in other parts of the turbine. 

Explanation. — The weight, E, flies out at the speed at which the 
emergency governor is set to operate and trips the trigger, T, This 
allows the spring, S^ to force the lever, L, free of the set screw, C. The 
steam in the pipe, P, then raises the valve, F, and escapes so that the 
pressure in P falls. The steam is thus allowed to escape from opening 
J (Fig. 151) of the automatic throttle valve and from the opening (F, 
Fig. 149) of the safety governor valve, so that both the throttle and the 
governor valves are closed (see Sec. 138 and caption to Fig. 151) whereby 
the steam supply to the turbine is cut off. 

Sec. 143] 



Fig. 151. — Westinghouse automatic throttle valve which is used in connection with 
the safety stop or emergency governor of Fig. 150. (So long as the emergency governor 
does not release the pressure at J, the valve may be operated as a common throttle 
valve. The pilot valve, A, and cylinder, C, balance the valve to assist in opening. 
The spring, P, prevents chattering. When pressure is released at /, the trip piston, L, 
moves due to the live-steam pressure behind it, and trips the lever, T, allowing the sleeve, 
V, to fall. The dash-pot spring, M, then closes the valve. Too rapid movement of the 
valve is prevented by the oil dash-pot and plunger D. The valve may be re-set by 
turning the hand wheel at its top until the sleeve, V, is lifted suflBciently that the trip 
lever, T, may be put in place.) 


144. A General Electric Co. Multi-ported Valve Governor 

is shown in Fig. 152. The steam for the turbine is admitted 
to the space, S, through the strainer, T. There are shoulders 
(not shown) on the valve stem which are so arranged that, 

i'Sfeam Exhaust 
/Relay Val/e 
'Steam Inlet 


Fig. 152. — ^Section of multi-ported governor valve used on some General Electric Co. 

Curtis turbines. 

as the valve stem lifts, the valves, A, B, C, and D are opened 
successively so that only one valve is opening at a time. The 
rest are all either closed or open. The various valves admit 
steam to the various nozzle passages, N. Thus there is very 
little throttling action and the governing is accomplished 

Sec. 145] 



■Cam Shaft 

chiefly by varying the number of nozzles to which steam is 

145. A General Electric Co. 
Multiple -valve Governor Mech- 
anism is shown in Fig. 153; this 
figure shows in section one of a 
number of similar valves which 
are arranged side by side along 
the top of the turbine casing. 
The governor proper (shown in 
Fig. 154) operates an oil-relay 
valve (F, Fig. 155) which admits 
oil against an operating piston. 
This piston moves a rack, R, 
which engages a pinion, L, on 
the shaft {S, Figs. 153 and 155). 
On this shaft are a number of 
cams, C, keyed at different angles. 
Thus when the operating piston 
moves, the cams strike successively 
their cam -folio wing rollers, R, and 
lift the various poppet valves, F, 
in turn. These valves admit 
steam to the various nozzles and 
bypasses of the turbine. 

146. Speed Adjustments Of 
One Or Two Per Cent. In Spring- 
opposed Governors such as that 
shown in Fig. 154 {General Electric 
Co. Instruction Book No. 82,207) 
may be made by varying the ten- 
sion on an external spring. This 
governor is used with the relay 
valve of Fig. 155 and the valve 
gear of Fig. 153. Governors of 
this sort are provided with aux- 
iliary springs, A, for varying the 

speed in synchronizing. If it is desired for any reason to 
permanently change the speed at which the governor operates, 


Fig. 153. — Controlling valve used 
for some General Electric Co. Curtis 
turbines. These valves are ' con- 
trolled by the governor proper shown 
in Fig. 154 through an oil-relay valve 
and rack-and-pinion device. 


Auxiliary Of 
Synch ron iz in^ 
^^ S prlncf^ 






tiynzhronizincf Motor;' 

Fig. 154. — Vertical centrifugal governor used on large-capacity General Electric Co. 

Curtis turbines. 


It-Retc«y Valve De+e\il 

Fig. 155. — Rack-and-pinion mechanism and hydraulic cylinder used for operating 
governor cams on large General Electric Co. Curtis turbines. 


this should be done by adjusting the nut, N, on the top of 
the governor. Adjusting N will, without affecting the speed 
regulation, change the speed only through a comparatively 
small range, on either side of that speed at which the gover- 
nor was designed to operate. Too much adjustment of N 
will affect the speed regulation. If it is necessary at any 
time to increase or decrease the speed regulation of the gover- 
nor, this can, within very narrow limits, be accomplished by 
inserting more lead — adding weight — in pockets (not shown) 
in the weights, W, to diminish the regulation. To increase the 
regulation, take lead out. However, if a considerable increase 
or decrease in regulation is required, it should be secured by 
respectively decreasing or increasing the number of working 
coils in the main spring, *S^, by screwing the top spring plug, P, 
in or out. A quarter turn of the plug will effect a material 
change in the speed regulation. 

Note. — The Positive Action Of Ant Governor Is Necessarily 
Dependent Upon The Absence Of Friction From Its Moving Parts. 
All knife edges, K, (Fig. 154) and joints should, if wear causes any 
appreciable deterioration, be renewed. In order that wear may be mini- 
mized, the governor should be assembled in such a manner that all of its 
rotating parts run as nearly concentric as is possible. 

147. A General Electric Co. Governor Proper Which 
Employs Inertia And Centrifugal Force As Governing Forces 

is shown in Fig. 156. The two inertia arms, A, carry the 
centrifugal weights, W, and the inertia weights, I. As the 
speed increases the centrifugal weights fly out against the ten- 
sion of the spring. The arms are affected by inertia and 
prevent sudden change in speed. The horizontal movement 
of the arms is changed to a vertical movement by two toggle 
levers, T, which fit into ball sockets on the arms. 

Note. — To Increase The Turbine Speed With This Governor 
(Fig. 156) without changing the speed regulation, subtract weight from 
the weight socket, W, or vice versa. The weight of opposite weights, 
W, must be kept equal to prevent unbalancing the governor. Increasing 
the main spring tension increases the speed and also decreases the speed 
regulation, and vice versa. Shortening xtie effective spring length by 
screwing the plugs, P, closer together increases the speed regulation, and 


vice versa. The governor is adjusted at the factory and need not, 
ordinarily, be altered except by the external adjustment (not shown) 
which is provided for the purpose. 



'r5p indie 

H- 5 e c + i o n x-X 

Fig. 156. — Inertia governor used on medium-capacity Curtis turbines. {General 
Electric Co. Bulletin.) 

148. An Air-pressure Or Pneumatic Governor Used On 
The Ridgway Steam Turbine (Fig. 157) employs an air- 
pressure blower, B, (directly connected to the shaft to furnish 
the operating power for the governor) instead of employing 

Sec. 148] 



the centrifugal force developed by weights as do most governors. 
The blower creates an air pressure which is approximately 
proportional to the square of the speed. This pressure is 

exerted on the under sides of two light aluminum pistons, P, 
the movement of which is opposed by a spring, S. The ten- 
sion on this spring is varied by the handwheel, K, or by the 
synchronizing motor, L. The double beat throttle valve, V, 
is controlled by the operating piston, D, through the oil-relay 



valve, G, and the floating lever, J?, in the usual manner. 
There is a spring, A^, which closes the valve in case of failure 
of the oil pressure. The chief advantages claimed for this 
method of governing are simplicity and absence of any high- 
speed parts on which there is friction. The runner of the 
blower has no friction except that of the air. 

149. A So-called Mechanical Indirect Centrifugal And 
Inertia Governor Valve Gear (Figs. 158 and 159) is in use on 

some medium-capacity (say 
500 kw.) General Electric 
Co. Curtis turbines. The 
illustrations show only one 
valve mechanism; on a tur- 
bine there are a number of 
duplicate mechanisms 
mounted side by ^ide, all 
controlled by a single gover, 
nor and each admitting steam 
to or cutting it off from one 
nozzle section. This valve 
gear operates (see explanation 
below) in a way somewhat 
analogous to a detaching Cor- 
liss-valve mechanism for 
steam engines. That is, it 
employs two pawls or '^pick-up 
hooks," A, for each valve. 
The pawls are attached to 
K and are oscillated up and 
down by the motion trans- 
mitted to K by L. The upper hook, A„, opens and the lower 
hook, Ac, closes the valve. The position of the shield plate 
or ''knock-off cam," E^ is controlled by the governor and 
determines the height to which the valves, Y (Fig. 159), are 
lifted. Unlike the Corliss mechanism, however, each valve is 
closed by a pawl, Ac, instead of being closed by springs or 

Explanation. — The lever, K (Figs. 158 and 159), is oscillated up and 
down by an excentric and the rod, L, at the rate of 120 complete 
strokes per minute. The pawls, A, are pivoted at P and P on the 

Of Cross head 

J Upper Position 
^•.,0f Crosshead 

-■Valve Stem 

Fig. 158. — Lifting and knock-off mech- 
anism of the Rice mechanisal valve gear. 
{General Electric Co.) 

Sec. 150] 



lever, K. Due to the tension of springs, S, on lugs, F, the pawls 
tend to engage the latch blocks, B, so as to carry the governor 
valves, V, up and down also. But the position of the shield plate, E, is 
controlled by the governor. It allows the governor valves to be lifted 
when the turbine requires more steam. Also when the turbine requires 
more steam, it prevents the valves from being closed on the return 
stroke. When less steam is required, the shield plate is so moved by 

Ovsshead- . . p 

Shield Plafe 

Fig. 159. 

-Rice mechanical valve gear used on some medium-sized General Electric Co. 
Curtis turbines. 

the governor as to allow the governor valves to be closed and it similarly 
prevents them from being opened. 

150. Dash-pots (Z), Fig. 159) are used on many turbine 
governors to prevent hunting. If a large centrifugal governor 
were so adjusted as to allow a regulation of only 1 to IJ^ per 
cent, in the speed of the turbine, the governor would have a 
tendency to vibrate slowly — or to move above and then below 


its correct position. A dash-pot is therefore frequently used 
to ''dampen" such vibrations of the governor and to main- 
tain it in its correct position. 

Note. — To Make A Governok More Sluggish, or slow-moving, use 
a heavier oil in the dash-pot, or restrict the opening around the plunger. 
To make it more prompt, give it more opening or thin the oil with 

Note. — For more complete directions for the care, construction and 
adjustment of dash-pots, see the author's Steam-engine Principles 
And Practice. 

151. In Adjusting A Governor To Synchronize Steam 
Turbo -alternators, a motor-operated device which is con- 

.-Synchronixlng Spring 

.,'Goyernor Ley^r 

Operating pjj^f 
Cylinder^, yalre'\ 


Fig. 160. — A General Electric Co. synchronizing device for turbo-generators which 
may be controlled from the switchboard. When the motor, M, is connected into circuit 
at the switchboard, it turns worm, W, and tightens or loosens synchronizing spring, S, 
depending on the direction in which the motor is caused to rotate. 

trolled by the switch-board operator is often employed. If a 
turbo-alternator is to be connected in parallel with another 
which is already running under load, it is necessary 
that, at the instant of connecting the one in: (1) The two 
machines he running at exactly synchronous speed. (2) The 
two machines he delivering the same voltage^ as shown hy a 


voltmeter. (3) The two machines he in phase. The ''dead" 
machine, which is to be connected in, is usually synchronized 
with the ''live" machine, which is already under load, by 
altering the speed of the dead machine until its speed is 
exactly the same as that of the live machine and the two are 
in phase. 

Note. — To Adjust The Speed For Synchronizing: On the smaller 
turbines, this may be effected by hand adjustment of the speed — changing 
the synchronizing spring {d, Fig. 147 and S, Fig. 157). On the larger 
turbines, the speed alteration is accomplished by a motor-controlled 
synchronizing device (Figs. 157 and 160) which forms part of the gov- 
ernor. The synchronizing motor may, in order to change the speed, 
vary the tension of the governor synchronizing spring as in Figs. 157 
and 160 or it may change the position of the pilot valve with respect to 
the governor valve as in Fig. 145. In Fig. 145 this is effected by turning 
H. H may, if desired, be motor controlled. After the two machines 
have been synchronized and are operating in parallel the proper division 
of the load between them is accomplished by adjusting their governors, 
and adjusting the field rheostats to minimize the cross currents. Divi- 
sion of load cannot be accomplished with only the field rheostats; see 
the author's American Electricians' Handbook. The machine which 
is to pull most of the load must be given proportionally more steam. 

152. The Care Of Governors seldom includes anything 
more than oiling and occasionally re-packing a stuffing box or 
regrinding a valve. The operation of the governor should be 
examined frequently. On small turbines, the whole governor 
may be moved by hand to see that it moves freelj^ and shuts 
off the steam. If undue lost motion develops, or if any part 
of the mechanism shows undue friction, the difficulty should 
be promptly remedied as explained in Sec. 132. There is 
some simple method of making a small change in speed on 
nearly all governors; and sometimes adjustable weights are 
provided to change the regulation as in Sec. 147. But the 
manufacturer should be consulted before any extensive or 
radical adjustments are made. After any governor adjust- 
ment, the action of the device throughout its range should be 
noted to make sure that it is safe. 

Note. — The Elaborate Relay Governing Mechanisms Employed 
On Large Turbines Are Too Involved And Various To Admit Of 
Special Directions For The Care Of All Of Them. In general, 


there should be means of ascertaining at all times if the relay system is 
properly filled at the proper pressure with the operating fluid (usually 
oil); see Div. 10. There is, usually on large turbines, an emergency oil 
pump (Sec. 197) which will keep up the pressure in the oil system if the 
regular pump becomes inoperative. The governor proper of a relay 
governor operates exactly as do other spring-loaded fly-ball governors. 

■^•- 0/7 Return 
Fig. 161. — Illustrating the lubrication of a General Electric Co. Curtis turbine governor. 

153. The Emergency Governor Should, Preferably, Be 
Tested Daily by carefully overspeeding the turbine up to the 

Sec. 154] 



speed at which the emergency governor should operate. 
When thus testing, the 
speed, as indicated by a 
tachometer, should be care- 
fully watched. It should 
never be assumed that the 
emergency governor is un- 
necessary because the speed 
governor functions prop- 
erly. Additional protec- 
tion against overspeed is 

Note. — The Parts Of A 
Steam Turbine Governor 
Which Require The Most 
Oiling are the worm gears and 
thrust bearings. These are 
sometimes provided with 
forced-feed oil systems as in 
Fig. 161; see also Div. 10. It 
is very important that the hnk- 
age pivots be kept oiled and 
not be allowed to stick but as 
these move but little, they do 
not require much oil. 

154. The Principal Kinds 
Of Valves Used In Connec- 
tion With Steam Turbines 

are: (1) Throttle valves (Fig. 
162) which are used for 
admitting steam by hand 
to the turbine. (2) Safety- 
stop or emergency valves 
(Fig. 137) which are oper- 
ated by the emergency 
governor, sometimes the 
emergency governor trips 
the throttle valve. (3) 
Governor valves (Figs. 129 and 152) which are operated by 

Fig. 162. — Throttle valve with safety-stop 
attachment used on some General Electric Co. 
Curtis turbines. (Many are in use but they 
are now applied to new machines only in 
special cases.) 


the speed governor. (4) Nozzle valves (Fig. 163) which are 
used principally on small turbines for admitting steam to 
additional nozzles for heavy loads. (5) Bypass or stage valves. 

' ' -Nozzle -^yalye Point 
Fig. 163. — De Laval nozzle and valve. 


{Vi, Fig. 78) which are used for admitting steam to later stages 
of a multi-stage turbine to carry overloads; these valves may 

be operated by hand or by the 
speed governor. (6) Relief valves 
(Fig. 164 and Fig. 29) which are 
safety valves placed in the turbine 
casing to protect it against exces- 
sive pressures. (7) Atmospheric 
relief valves (Fig. 185) which allow 
the turbine to exhaust to the 
atmosphere if the condenser fails 
and thereby prevent the building 
up of an excessive pressure in the 
turbine casing; such valves are con- 
nected as side outlets in the exhaust 
pipe between the turbine and the 
condenser. See the author's Steam 
Power Plant Auxiliaries And 


Pipe thread 

Fig. 164. — A relief valve suit- 
able for use on a steam turbine. 
{Ashton Valve Co.) 

Note. — Throttle Valves For Small 
Turbines are usually ordinary globe valves in the steam pipe near the 
turbine. For larger turbines, the throttle valves are more elaborate as 
shown in Figs. 151 and 162, and act also as safety-stop valves. The 


balancing pistons of these valves are subject to some of the troubles 
of engine pistons, although a certain amount of leakage past these 
pistons is expected. 

Note. — A Sentinel Valve {Kerr Turbine Co.) is a valve which is so 
placed and designed as to allow escape of steam and thereby give warning 
if the pressure becomes high in the low-pressure end of the turbine casing. 
Overload valves are valves which are opened to carry overloads, that is to 
give the turbine more power than its normal rating. They are, ordi- 
narily, stage valves or nozzle valves and may be operated either by hand 
(for small turbines, usually) or by the speed governor (for large turbines) 
depending on the construction employed. 

155. The Chief Troubles With Valves Are; {1) Stuffing-box 
leaks; (2) Valve leaks or breaks; (3) Sticking. Stuffing-boxes 
can be repacked with various types of high-temperature pack- 
ings which are on the market for the purpose. For most 
saturated-steam valve stems, candle-wicking soaked in oil 
may be used. A governor-valve stem must be packed very 
carefully so that it will hold steam without much friction of 
the packing. It is usually better to first screw the gland nut 
up tightly and then slack it off so as to relieve the pressure 
on the stem. In general, it is better to have a slight steam 
leak around a governor-valve stem than to have too much fric- 
tion. Some indications of a leaky governor valve are: (1) 
Racing at light loads with the valve apparently closed and (2) 
heating of the governor thrust hearing due to the force developed 
by the governor in endeavoring to close a leaky valve. One test 
for valve tightness is to close the valve by hand while the 
turbine is running and note how rapidly its speed decreases. 

Note. — Common Causes Of Governor Valve Failure are wet steam 
and running constantly at light loads. Wet steam may be avoided by lag- 
ging the steam pipes and installing a separator. Running at light loads 
will not wear the valve if one or more of the nozzle valves are turned off. 
If this cannot be done, a smaller valve should be used. It is necessary to 
ascertain from the manufacturer what is the smallest valve which will 
carry the required load. If a conical-seated valve is reground occasion- 
ally, it may be kept in good condition in spite of continued running at 
light loads. 

156. Steam Strainers (Figs. 152 and 165) are provided in the 
admission passages of most steam turbines. They are usually 
located so that the steam is strained before it passes the gover- 


nor valve. This is a precaution to prevent particles of scale 
from the pipe and other foreign matter from getting under 
the governor valve and preventing its shutting. Strainers 
are commonly constructed of sheet metal in which holes are 
punched which are sufficiently large to allow the necessary flow 
of steam but small enough to keep out any solid particle which 
would damage the turbine. The total area of the holes is 



Punched S/of3. "'• 

Supporting Ring..-' 
Fig. 165. — Common type of steam strainer used on small turbines. 

generally made much larger than that of the rest of the preced- 
ing and following passages so that there will not be much 
friction in the strainer. 


1. Under what conditions may a turbine be operated without a governor? Why is 
a governor usually necessary? 

2. Show by a sketch the action of an elementary direct throttling governor. 

3. Of what principal parts does a complete governing mechanism for a large turbine 
ordinarily consist? 

4. In what three ways may steam turbine governors be classified? Name at least 
two subclasses under each classification. 

5. What is one disadvantage of a throttling governor? One advantage? 

6. Explain the use of a block or key in adjusting the lost motion on a Sturtevant 

7. What is the function of a vacuum breaker on a governor? 

8. What method may be used for making speed adjustments of about 2 or 3 per 
cent, on nearly all horizontal throttling governors? What should be done in case it is 
desired to make a radical speed adjustment on a governor? 

9. How may corrosion of governor valves be minimized? 

10. What is an emergency governor? Show by a sketch how a simple one may 

11. At about how much greater than normal speed is the emergency governor usually 


12. What is the function of an oil-relay mechanism for a steam-turbine governor? 
Draw a sketch of and explain the operation of such a mechanism. 

13. What is the advantage of an intermittent governor over a throttling governor? 

14. Explain, using a sketch, the action of a floating lever in a relay governor. 

15. What is the effect of decreasing the number of coils of a governor main spring? 

16. How does the Westinghouse safety stop control the automatic throttle valve? 
Use a sketch in explaining. 

17. What is the purpose of bypassing in a multi-stage turbine? What are its dis- 

18. How do multi-ported governor valves avoid loss of energy by throttling? 

19. What is the function of an inertia arm in a governor? 

20. Explain the operation of a pneumatic turbine governor. What are its advant- 

21. To what steam engine mechanism may the Rice mechanical valve gear be 
compared? Explain the Rice governor using a sketch. 

22. What is the function of a dash-pot on a governor? How may the piston on one 
be made to move more slowly? More rapidly? 

23. How are turbo-alternators usually synchronized from the switchboard? 

24. What is a throttle valve? How may it be interconnected with an emergency 
governor? Explain with a sketch. 

25. What are bypass valves? Atmospheric relief valves? 

26. What is a sentinel valve? A relay valve? 

27. What are the three chief troubles encountered in valves? 

28. How may leakage in a governor valve be detected? How repaired? 

29. What steam and load conditions tend to wear out the valves of throttling 

30. What is the general construction of most steam strainers for turbines? What is 
their function? 



157. The Function Of A Steam-turbine Reduction Gear 

(Fig. 166) is solely to "reduce" the rotative speed of the tur- 
bine shaft to a suitable speed for driving some other machine. 

Fig. 166. — A single-stage Moore steam turbine, showing the method of mounting 
turbine and reduction gears on a common bedplate. 

Since turbines can be operated efficiently only at high rotative 
speeds (see Div. 3) and since many mechanically driven 
machines must be operated at low rotative speeds, it is 
obvious that these low-speed driven machines cannot be 
coupled directly to the turbines. Strictly speaking, a reduc- 
tion gear does not reduce the speed of the turbine shaft. 
Rather, the turbine shaft transmits its power through the 




reduction gear (or gears) to another shaft which then is 
connected to the driven machines. 

Note. — Reduction Gears Are Often Not Necessary With the 
following machines : (1) Alternating-current generators. (2) Small direct- 
current generators (below about 50 kw.). (3) Centrifugal pumps. (4) 
Fan hloivers. (5) Turbo-co7npressors. Nearly all other machines must 
be driven at much lower speeds than those at which steam turbines 
operate and, hence, require reduction gears, 

158. Steam-turbine Reduction Gears May Be Classified 

as follows: (1) Single-reduction gears, Fig. 166. (2) Double- 

Firsf Second ^ 

.Rec/ucfion ^'9^''^. ,.^ Redact ,on.^ 

^..•■Coupling ; 

Fig. 167. — Single-plane-tj-pe, double-reduction gears for a 3,000-hp. marine turbine 
which reduces the speed from 3,500 r.p.m. at the turbine to 90 r.p.m. at the propeller. 
(De Laval.) 

reduction gears, Fig. 167. (3) EpicycUc gears, Fig. 171. 
Single-reduction gears may be employed whenever the turbine 
speed does not exceed about six or eight times the speed of the 
driven machine. Double-reduction gears are employed for 
greater speed reductions than can be accomplished with a 
single reduction. By employing a double reduction the sizes 
of the gears may be kept smaller than if the total reduction 
were accomplished with one gear and one pinion. The epi- 
cyclic-gear speed reducer is explained and discussed in Sec. 162. 


Note. — Double-reduction Gears Are Used Extensively With 
Marine Turbines and occasionally for such stationary service as mill 
or shaft drives. Double-reduction gears whose shafts all lie in one plane 

j3llddoj(j qi- - 




A % 



X> +j 






fl o 

^ S- 






c o 

C « 


o g 





i/^j/^/^/ uoipnpdy fsi- 

(Fig. 167) are called single-plane gears to distinguish them from those in 
which the driven shaft is located lower than the turbine shaft. Two- 
plane gears (Fig. 168) usually transmit power from two turbines to a 
single slow-speed shaft. 



159. The Construction Of Reduction Gears is usually such 
that the gears are enclosed in a case (Fig. 169) which serves 
to exclude dust and other foreign matter from the teeth. 
The gears are usually cut from high-grade rolled steel. The 
teeth are of the double-helical or herringbone type and thus 

2^--Liftin0 Eye 


.'Supply Line To Of I Cooler Case Cap-. 
\ OearBearing, /^ 

I Gear thrive / Q'f 

' T fhmp 

•Supply Line ToBearings 
From Oil Cooler 

Supply Line 
• To Oil Spray 
\ Tube 
^ Oil- Pump Coupling 


~- Overflow 
From Bearings 


^- - Gear Case 

'Spray Tube For Oiling Gears 

Fig. 169. — Side sectional view of double helical reduction gears. (Moore Steam Tur- 
bine Corporation, Wellsville, New York. Instruction Card No. 4.) Forced-feed lubrica- 
tion is used in all Moore reduction-gear sets. The oil is supplied from a geared pump, 
P, under pressure, to the bearings, B, and also is sprayed through small holes in a copper 
pipe, T, onto the gear, G, and pinion, iV, at the pitch line. Stop cocks are provided in 
the feed lines to the bearings for regulating the flow and also in the supply line for spray- 
ing oil onto the gears. These cocks should be adjusted so both bearings and gears will 
receive a liberal supply of oil. The bearings should be given all they will take without 

Inspection of gear lubrication can be made through the opening (not shown) which is 
provided for this purpose. A metallic ringing sound is an indication that the gears 
are not getting sufficient oil. If for any reason too much oil is fed to the bearings and 
gears, so that it is not carried away fast enough through the drain pipe and that it 
backs up in the case until the gear dips in the oil, there will result undue heating, caused 
by the oil being thrown against the sides of the case. The remedy is to reduce the quan- 
tity of oil which is used. 

A cooling device is provided in the form of a brass-tube cooler or plate-type cooler for 
cooling the oil. Water is used for cooling. The oil is circulated from the discharge 
of the oil pump through the brass-tube cooler. In the plate cooler, the oil passes over 
the cooling surface when it is being returned to the suction tank. 

provide smooth quiet operation which is free from vibration 
and end thrust. The gears are supplied with oil, as are also 
the bearings, from a pump (see Div. 10) which is driven from 
the end of the large-gear shaft. The oil is cooled by passing 
it over a water-cooled coil or plate and is then returned to the 
pump. Some large marine reduction gears are so designed 


that the pinion shaft turns in a floating frame carried on 
hydrauhc rams. This elastic support of the pinion shaft 
renders the gearing practically noiseless and insures automat- 
ically more nearly perfect alignment between gear and pinion 
under all conditions. With turbines of smaller output, how- 
ever, the floating frame is seldom used. 

Note. — The Transmission Efficiency Of Reduction Gearing is 
very high; it may exceed 98 per cent. The transmission efficiency = 
{the power delivered at the low-speed shaft) -^ {the poiver developed by the 
turbine). This efficiency is materially decreased, however, if the oil level 
is permitted to reach such a height that the gear dips into it or if too little 
oil is supplied to the gears. 

160. Troubles With Reduction Gears are infrequent. The 
principal care which reduction gears require is to see that they 
are maintained in proper alignment and that they are properly 
lubricated. Misalignment causes vibration and rapid wear 
and is frequently the cause of noisy operation. When lining 
up the gears bear in mind that either the gear or the pinion, 
depending on the direction of rotation, will be lifted to the 
top of its bearings when the gears operate. When the gears 
run toward each other at the top the pinion will lift. When 
the gears run away from each other at the top the gear will 
lift. Note the clearance in the bearings by lifting on the 
shaft; the clearance is the amount of ''give" of the shaft in 
the bearings. Then make adjustment for about 0.002 in. less 
than the observed clearance. For the lubrication of high-speed 
reduction gears a good gear oil should be used. See Div. 10. 
The oil should be kept clean by renewing or filtering it as 
often as is found necessary. The temperature of the oil 
should be maintained at between 130° and 180° F. 

Note. — The Oil-cooling Coils Of Reduction Gears should be sup- 
plied with cool clean water in sufficient quantity that the oil is kept at 
the proper temperature (see above). The water piping should be 
arranged that the coils may be protected against possible freezing. 

161. The Alignment Of Reduction Gear, as given by the 
Westinghouse Electric and Mfg. Co. in **Instruction Book 
No. 5,220" is as follows: 



1. Alignment In A Horizontal Plane. — Check the alignment with 
the block gages furnished. If these are unavail'able, caliper between the 
aligning collars (C, Fig. 170) on pinion and gear wheel and note the 
micrometer measurement. Micrometer the gear and the pinion-aligning 
collars. The center to center distance between gear wheel and pinion 
shafts is thus determined by calculation from these measurements on 
each side of the gear wheel, and should of course be the same, within 
0.001 in. If it is not, shift a liner from the proper side pad of the pinion 
bearing, B, to the opposite side. A 0.000,5-in. Hner will affect the differ- 
ence in center-distance dimensions about 0.001 in. Adjust at the pinion 
bearing, B, in preference to the turbine bearing, A, since the latter 
throws the glands slightly more out of center. 

2. Alignment In The Vertical Plane. — This alignment can be 

ferm/ha/ /^t/^^^ ^ Aligning. ^ ^ 

diock ' - . Bearing ^^ <Co//ars ^ (Cd 

~«, Pinion-^ \ 


I Bearing ^ 

Turbine-Wheel • 

Fig. 170. — Small (15-50 kw.) geared-turbine and generator. (^Westinghouse Electric 

& Mfg. Co.) 

properly checked only by the operation of the unit. As a rough approxi- 
mation, coat a few pinion teeth with Prussian blue and pull the turbine 
rotor around in the direction of its rotation. Then note the distribution 
of the contact marks on the gear teeth. If these seem to be concentrated 
at the ends of the teeth, say at the turbine end of each helix, raise the 
pinion bearing, B, by shifting a liner from the top pad to the bottom 
one and repeat till the contact appears distributed rather than concen- 
trated. This is not a complete check, since, under load, the pinion takes 
a slight deflection. To thoroughly check, prepare the gear-wheel teeth 
by washing them with a copper-sulphate solution, thus giving a light 
film of copper deposit which will plainly show the contact of the teeth 
■during operation. When everything else about the unit is ready, run 
the turbine for half an hour under approximately full load. Then remove 
the gear case cover and examine the contact marks on the gear teeth. 


These should extend from end to end of the teeth. If the marks are 
concentrated at either end, transfer a 0.005-in. liner as directed above, 
again apply the copper-sulphate solution and repeat the trial run. Closer 
pad adjustment than 0.005 in. is not required, even though the tooth 
contact marks might seem to indicate it. 

3. Backlash Or Clearance Of Teeth. — Block the gearwheel 
against end movement. Push rotor and pinion to one end as far as 
possible and take a feeler-gage measurement where convenient, say 

5f a Nonary Gecr __ 

(Does iVrf Revo/ve) -^C-^^^att^ 

Conf^ecteof 7o 

■ Pfanchirtj 

Or fpicyclfc 
Gear^; Sh'ifts 

\ Are Mo ur -I ted 
In A Cage Whic / 
Revolves Af lov\ 

Fig. 171. — Illustrating the principle of the "Turbo-Gear" speed reducer: Annular 
gear Gz is so held in the frame of the unit that it cannot revolve; pinion shaft 5t is re- 
volved at high speed. (Epicyclic reducing gears as manufactured by the Poole Engineer- 
ing and Machine Company.) 

between the gland runner and casing. Pull the rotor in the opposite direc- 
tion and again take a feeler measurement. The difference, or end play 
of the pinion should be between 0.009 and 0.016 in. In taking such 
measurements be sure that glands or blades do not strike adjacent parts, 
thus giving false values. If necessary to correct end play, alter center 
distance by shifting equivalent liners of both turbine and pinion bearings, 
A and B, from one side to the other. Operation (1) has already put the 
shafts parallel and therefore one bearing should not be changed without 
changing the other the same amount. The end play will be changed 
about 0.005 in. by shifting a 0.005 liner. 

162. Epicyclic Reducing Gears (Fig. 171) are so formed 
that, although they afford but a single reduction, the driven 



or low-speed shaft has its axis exactly in line with the driving 
or high-speed shaft. They are installed in a frame (Fig. 
172) which presents the same general appearance as an 
enclosed electric motor or generator. Under certain condi- 
tions their construction makes them more applicable than 
ordinary single-reduction gears. Their operation is obvious 
from Fig. 171. 

Planetary Gea/rG^ 
Casing, F- 

Infernal Gear, G3 

Pilot Bearing 
-•Inspection Pane/hole 

Main Bearing, 

Shaft •; 

■Oil outlet 
to Bearings^ 




Fig. 172. — Longitudinal section through the Turho-gear speed reducer, Fig. 171. The 
low-speed shaft, Sj)^ carries the cage, E, and is supported in the casing, F, on the two 
ball bearings B and C. The cage, E, contains 3 pins, P, upon which the planetary 
gears G2 revolve. The pinion shaft, S^^ carries the pinion, Gi, which meshes with the 
three planetary gears, G2; <Sy is carried in the two bearings A and K. The planetary 
gears Gz "roll around" in the internal gear Gz, which is held stationary — so that it can- 
not turn — in the casing, F. An oil pump is driven by the eccentric on the low-speed 
shaft, Sjy. 

163. Steam-turbine Couplings Are Of Two Kinds: (1) 

Rigid, (Fig. 173) (2) Flexible (Fig. 174) see Sec. 164. Rigid 
couplings are employed principally on small turbines and 
only where both the coupled turbine and driven shaft are 
supported on only two or three bearings. Where four bear- 
ings are used, two for the turbine shaft and two for the driven 
shaft, a flexible coupling (Sec. 165) is always employed. 

Note. — The Rigid-coupling Two-bearing Unit Is Very Desirable 
for small-power machines. There is much less chance of such a machine 


getting out of alignment and thus giving bearing trouble. A two-bearing 
unit also occupied less floor space than does a three- or four-bearing unit. 
The Steam Motors Co. of Springfield, Mass. specializes in two-bearing 
units which it builds in sizes up to 300 hp. 

''Labyrinth Gland ' 


Outboard- beanncf 

Fig. 173. 

Assembled rotor for a "Steam Motor" generator-set showing rigid flange 
coupling. {The Steam Motors Co.) 

1 Rubber 
' Bushing 

164. The Purpose Of Flexible Couplings In Steam-turbine 
Drives is: (1) To provide for any slight inequality in the wear 
of the hearings. (2) To permit axial adjustment of the turbine 

spindle. (3) To allow for differ- 
ences in expansion. It is obvious 
that two shafts, each supported 
on two bearings, would be bent 
by any deviation of their bear- 
ings from one straight line. 
Furthermore, it is very difficult 
to exactly align four bearings 
into a straight line and, if 
aligned, to so maintain them. 
Hence, and to permit of axial 
adjustment of the two coupled 
shafts, a so-called flexible coup- 
ling (Fig. 175) is employed; see 
Sec. 165. 

165. There Are Three Princi- 
pal Types Of Flexible Couplings, 
namely: (1) Ruhher-hushing type, (Fig. 174), wherein a number of 
— usually six — coupling bolts or pins are fastened rigidly to 


Turbine Coupling 

■These Faces Nusf 
This Dimension '■' Be Parallel At 

Should Not Be Less „ All Points 

Than^" Nor More Thani 

Fig. 174. — Final alignment of 
Type-6 Sturtevant turbine coupling. 
The turbine rises when steam is turned 
on. Therefore provide allowances to 
compensate for this change. It is 
important that final alignment be 
made under operating temperatures. 



one half of the coupling and are extended through rubber (or 
leather) bushings in the other coupling half. The rubber 
affords the flexibility. (2) Flexible-pin type, (Fig. 175) wherein 
flexibility is attained through the bending of small driving pins 

/Cap Refains ^p^rce Here Provides Flexibi/ifu. 
\ B ushing ^ , -■■ 

. ' VrWing Pins-^ 

E-^ End Yiew_ 


dross 'dushing ' ^J 

1-Longitudinal Section 


I2-Sect\on B-B 

Fig. 175. — Pin-type flexible coupling used on Westinghouse turbines. 

P — this type employs no highly compressible material; some- 
times the pins, P, are built up of small sheet-steel laminations. 
(3) Claw or jaw type, (Fig. 176) wherein flexibihty is attained 

^^i^/;^lor ,. Coupling Ends- - . .^ 

;Oll Orer-Flow 

\^ Coupling Housing^ 

i ■ — > Oil HoIeS' , — -k-q „>, „ ^ 

>Mw^y^/y^y//y/y^M////AjJ/y ///M,w.vY'///7 ?7777^A "Oil Possage 

1-LongItudInal Section H-Troinsversc Section 

Fig. 176. — Claw-type flexible coupling used on* Allis-Chalmers turbines. 

through the joints between the coupling jaws and the claws 
on the sleeves. Couplings of types (2) and (3) require lubrica- 
tion of the driving surfaces because there is sure to be some 
sUding between the metal contact parts. 


Note. — The "Flexibility" Of A Flexible Coupling is very small; 
that is, a flexible coupling will permit of very little misalignment of the 
two shafts which it connects. Under operating conditions (turbine hot) 
there should not be over 0.002 in. difference in height between the two 
halves, nor should the angular misalignment between the connected 
shafts be such that the difference in opening between the two halves on 
opposite sides of the shaft exceeds 0.002 in.; (see Fig. 174) and Sec. 167. 
The principal mode in which a flexible coupling affords much "play" is 
in the axial direction. 

166. The Care Of Steam-turbine Couplings is simple. 
Rigid couplings, once installed, require no further care. The 
bolts must be so fastened, however, that they cannot come 
out — note the ''wire-lock" fastenings in Fig. 175. All-metal 
flexible couplings must always be lubricated. All flexible 
couplings should be examined periodically (say once a month) 
to see that the connected shafts have not become misaligned 
by wear or other causes. Should the couplings need aligning 
proceed as directed in Sec. 167. Coupling parts which, when 
an inspection is made, show considerable wear should be 
repaired or the worn parts replaced. 

Note. — Serious Misalignment Of Shafts Results In vibration, 
hurned-out bearings, broken shafts, broken couplings, or broken other rotating 

167. A Convenient Method Of Aligning Two Shafts At 
Their Coupling is given below. Two shafts may suffer from 
two kinds of misalignment. They may be out of line sideways 
(the ends of their axes not meeting) or they may be nonparallel. 
The following method of checking their alignment is simple, 
always applicable, and can be performed in a few minutes: 

Explanation. — With a pin-type coupling, insert a coupling pin, with- 
out its bushing, through both halves of the coupling and leave this in 
while measuring. During all of the following measurements see that the 
couplings are pushed as far apart as the thrust bearings will permit. Make 
two marks, X and Y, one On each coupling, as shown in Fig. 177. With 
these points up, as shown in Fig. 177, measure distance A using a feeler 
or thickness gage. Measure also distance B using a steel straightedge, 
as shown in Fig. 174, and a feeler gage. Record these distances as shown 
in Fig. 177. Then turn the points to the right-hand side and repeat the 
measurements at the marked points. Repeat the measurements with 
the points in the down- and left-hand positions. If all of the dimensions 



A are the same, the two shafts are parallel. If all of the measurements 
B are the same, then the two shafts are not out of line sideways. If both 
of these conditions are not fulfilled, the shafts should be adjusted by- 
shifting or shimming the bearing pedestals or linings until the shafts are 
perfectly aligned. 

With a claw-type coupling, a test rod, C (Fig. 178) should be clamped 


Itnportani No+e: 

T/ie Two Shafts Must Always 5e Turned Over 

Together While Measuring 5oThaf Points X And Y 

On Each Of The Couplings Are Always Opposite Each Other 


Di^+ance"A''At Points"X"And"Y"A5 
Shaf-t-5 Are Turned OverTogether 
To Varying Positions 

Dis+ance*B"A+ Points"X"And Y As 
Shafts Are Turned OverTogether 
To Varying Positions 









R.H. Side 


R.H. Side 






L.H. 5ide 


L.H. Side 


The Above Indicates That Shafts Are Pafallel And In Line. If 
Dimensions Are Not Constant For Every Position, Thien They 
Should be Made So by Shifting Or Shimming Under Feet Of 
One Of The Members. 

Fig. 177. — Example, illustrating method of aligning couplings. 

Test Rod--^r 

D ..-Wedges To Hold 
Test Rod 

Fig. 178. — Showing a method of aligning a claw coupling. 

in one of the coupling ends as at D. The distance between the other 
coupling-end and the point of C should be measured with a feeler gage 
as was distance B, Fig. 177. The distance between two claws directly 
opposite each other should be measured (with calipers or micrometer) 
in the same manner as was distance A, Fig. 177. For the shafts to be 
in line, these two distances must be the same for any position — wheth'^r 
the points measured are up, down, or on the right or left side. 



1. Explain the function and purpose of reduction gears. 

2. What classes of machines are frequently driven without reduction gears? Name 
some with which gears are necessary. 

3. Name and distinguish between the three principal types of reduction gears. 

4. What are the principal uses of single-reduction gears and give their limitations. 
6. What are the principal uses of double-reduction gears and what two types are 

there? What determines largely which type is required? 

6. Explain, with a sketch, the usual construction of reduction gears. Describe the 
floating frame construction. 

7. Define transmission efficiency. What is a usual value and what may lower it? 

8. Explain what care reduction gears require and what troubles must be guarded 

9. Describe fully, using sketches, the method of aligning the teeth of a pair of reduc- 
tion gears. 

10. Explain, with a sketch, the operation of epicyclic reduction gears. What lire 
their advantages? 

11. In general, what two types of couplings are employed on steam turbines? 

12. On what kinds of machines are rigid couplings employed? What are the advant- 
ages of such drives? 

13. Give three reasons for employing flexible couplings. 

14. Describe, using sketches, the three principal types of flexible couplings. Which 
types require lubrication? 

16. What can you say regarding the "flexibility" of the so-called flexible couplings? 

16. What care do steam-turbine couphngs require, if any? 

17. What harmful results are occasioned by poorly aligned turbine shafts? 

18. Explain, with sketches, methods of aligning pin and claw couplings. 



168. A Steam-turbine Regenerator Or Accumulator (Fig. 
179) consists of a large mass of water, W, which absorbs heat 
from exhaust steam when the steam is brought to it and which 
gives up heat by evaporation when required. A regenerator is 


From Boifer 



5feam From 


4-SidV Elcvrt+ion Showing Piplnoj Arrangement 

179. — A typical Rateau regenerator or accumulator for use with low-pressure 

generally necessary when exhaust steam from an intermit- 
tently used non-condensing steam engine, such as a rolUng- 
mill engine or a steam hammer, is used to drive a low-pressm-e 
turbine (Div. 9). A regenerator will insure a steady flow of 
steam to the turbine for a short time (about four minutes, 
usually) after the engine has been stopped. Regenerators 
should always be enclosed in an effective heat-insulating 



Explanation. — The regenerator of Fig. 179 consists of a shell, A, 
which is kept about two-thirds full of water and which contains two 
mixing tubes, B. Exhaust steam from the engine is led, first through 
an oil separator (not shown), and then through a check valve, F, into 
tubes, B. A slight steam pressure in B forces the water down in their 
vertical legs which have a large number of holes (usually about % in. in 
diameter) as shown. The steam then issues through these holes and 
bubbles upward or condenses, depending upon the temperature of the 
water, causing a circulation as shown by the arrows in II. The circula- 
tion is assisted by baffle plates, P. As soon as the water in, A reaches its 
boiling point, the space above the water level will fill with steam which 
then passes outward through cross, C, and pipe, T, to the low-pressure 
turbine. The baffle plate, D, prevents small drops of water from passing 
out through C. 

If, now, the turbine does not require as much low-pressure steam as is 
furnished by the engine exhaust, the steam will not be permitted by the 
turbine governor to flow through T. Hence, it will accumulate in A 
and raise the pressure. But, as the pressure in A increases, the boiling 
point of the water also increases. Hence the water will now absorb 
more heat. Thus more and more heat is stored in the water until, 
finally, the pressure in A reaches a value at which the back-pressure valve 
O is set to open. Then all steam which is not needed by the turbine will 
be discharged into the atmosphere. 

If, now, the engine should be stopped, will close and the turbine will 
draw steam from A. Thus the pressure in A will be gradually decreased. 
But, as the pressure is decreased, the boiling point of the water in A will 
be lowered and some of the water will be evaporated. Thus, steam will 
continue to be supplied to the turbine, but at a gradually decreasing 
pressure, until the engine is again started. Sometimes, however, a high- 
pressure steam pipe, S, is arranged with a reducing valve, R, to admit 
steam to C when the pressure in A falls below a predetermined value for 
which R is set. If the turbine is equipped with a bypass or high-pressure 
valve, the reducing valve, R, is not necessary at the regenerator. 

Note. — Piping Accessories Which Should Be Installed With A 
Regenerator are: (1) An oil separator; oil is generally undesirable in 
a steam turbine because it tends to adhere to the blading and clog the 
passages. (2) A check valve, V Fig. 179, to prevent water from passing 
from the regenerator back to the engine cyHnder when the engine is 
stopped. (3) A safety or hack-pressure valve, O Fig. 179, to prevent an 
excessive pressure in the regenerator which might be destructive to the 
turbine or the regenerator itself. (4) A float-valve water-level control, not 
shown in Fig. 179, to prevent an excessively high water level in the regen- 
erator; the water level will gradually rise as steam is condensed by the 
loss of heat from the regenerator shell by radiation. The water level 
may also rise because of the moisture which is carried into the regenerator 
with the exhaust steam. The water discharged by the float valve may 
be led to the hot well or permitted to flow into the sewer, whichever is 
most feasible. 



169. Regenerators Are Practical Only When the non- 
condensing engine which supplies the exhaust steam has short- 
period shut-downs. If the usual shutdown period exceeds 
three or four minutes, it is generally better to use a mixed- 
pressure turbine (Div. 9) than to attempt to use a regenerator. 
But in cases where the shutdown period seldom exceeds one 
or two minutes, a regenerator is very useful. 

Note. — Boiler-pressure Variations May Be Conducive To The 
Use Of A Regenerator. — \^Tien a large reciprocating engine is suddenly 
stopped, the boiler which supplied the engine continues to produce steam 
at the same rate as before. The steam pressure immediately increases 
and very soon the safety valves are blowing off steam. The regenerator 


High- Pressure 

High-Pressure 5feam For 
Auxiliary Supply To Turbine-. 



180. — Typical layout of a power plant with a non-condensing engine, E, regenera- 
tor, R, and low-pressure turbine, T . 

can be arranged to receive the steam, which would thus go to waste, in 
one of two ways: (1) The blowoff can be piped to the regenerator, R, Fig. 
180. (2) A relief valve may be provided to discharge steam from the boiler 
at 1 or 2 lb. per sq. in. less than that for which the safety valves are set, 
the discharged steam being piped to the regenerator. 

170. The Normal Operating Pressure For A Regenerator 

is generally between atmospheric pressure and 15 lb. per sq. in. 
gage. A small vacuum could be used but would make difficult 
the exclusion of air from the system. The relief or back- 
pressure valve (0, Fig. 179) should be set to open at about 2 lb. 
per sq. in. above the normal operating pressure. The reducing 
valve or regulator (R, Fig. 179) should be set for about 1 lb. 
per sq. in. below the operating pressure. Hence, the pressure 
variation in the regenerator should not exceed 3 lb. per sq. in. 
For economical operation neither the back-pressure valve, 0, 
nor the regulator, R, should open except when unusual condi- 


tions arise. This necessitates the use of a regenerator which 
contains the proper mass of water (Sec. 171). 

Note. — Adjustment Of Regenerator Operating Pressure should 
be so made, if possible, that the non-condensing engine will supply the 
same amount of steam as the low-pressure turbine uses. By increasing 
the regenerator pressure the non-condensing engine can be made to use 
more steam and the low-pressure turbine less. This, then, is the remedy 
when the -regenerator pressure always is low. By decreasing the regen- 
erator pressure the turbine can be made to use more steam and the non- 
condensing engine less. Obviously, when the regenerator pressure is 
always too high (indicated by blowing off), the blowoff valve should be 
set for a lower pressure. 

171. To Compute The Necessary Weight Of Water In A 
Regenerator, the following formulas may be used. For. (28) 
gives the weight necessary to insure that the regenerator 
pressure will not drop too low while the steam supply to it is 
cut off for a short time. For. (29) gives the weight necessary 
to insure that a sudden supply to it will not cause a discharge 
from the back-pressure valve. That formula which gives 
the greater weight should govern the installation. The 
formulas are: 

(28) w,.i= ^Y(r?i"rS'^ (P°™^') 

(29) W..= MLd_^ (pounds) 

Wherein: Wpri = the weight of water, in pounds, necessary 
to insure that the pressure will not fall below a predetermined 
point while the turbine is using steam but no steam is supplied 
to the regenerator. Wtf2 = the weight of water, in pounds, 
to absorb a momentary rush of steam, t = the maximum 
time, in minutes, during which steam is being taken from the 
regenerator while no steam is supplied to it. W^i = the 
total steam consumption of the turbine in pounds per minute. 
Ws2 = the weight in pounds of a momentary supply of steam 
which must be absorbed. Li and L2 = the latents heats of 
steam, in B.t.u. per pound, at the maximum and minimum 
pressures in the regenerator. Ti and T2 = the temperatures, 
in degrees Fahrenheit, at the maximum and minimum pres- 
sures in the regenerator. 


Example. — Determine the weight of water to be stored in a regenera- 
tor which operates a 1,000-hp. low-pressure turbine for 4 min. while no 
steam enters the regenerator. The regenerator pressure may vary 
between 20 and 17 lb. per sq. in. abs. The turbine uses 30 lb. of steam 
per hp-hr. Solution. — From steam tables, Ti = 228° F. T2 = 
219.4° F. Li = 960 B-t.u. per lb. L2 = 965.6 B.t.u. per lb. Hence, by 
For. (28), Wwi = ^Wsi(L, + L2)/2(T, - T2) = 4 X (1,000 X 30 h- 60) 
X(960 + 965.6) -^ [2 X (228 - 219.4)] = 223,900 lb. 

Example. — If the regenerator of the above problem is to absorb 3,000 
lb. of exhaust steam during a short period of sudden supply, how much 
water should it hold? Solution.— By For. (29), Ww2 = W82(Li + 
L2)/2{T, - T2) = 3,000 X (960 + 965.6) -r [2 X (228 - 219.4)] = 
335,850 lb. 

172. A Condenser, As Used In Connection With A Steam 
Turbine, is a vessel into which the exhaust steam from the 
turbine is led and wherein the steam is condensed into water 
or ''condensate." The purpose in so doing is to create 
as high a vacuum as possible in the chamber into which the 
turbine exhausts. The vacuum is formed by causing the 
steam to come into contact with cold surfaces, give up some 
of its heat, and thus change from the vapor to the liquid 
state. The degree of vacuum formed depends on how rapidly 
heat can be carried away from the steam. The effect of high 
vacuum is to greatly increase the amount of heat which is 
liberated by each pound of steam and which may be converted 
into work by the turbine. See Sec. 10 for methods of comput- 
ing the liberated heat at various vacua. See also Div. 13 for 
the effects of vacuum on steam-turbine economy. 

Explanation. — The turbine, T, (Fig. 181) exhausts steam at S. This 
steam comes immediately into contact with the tubes inside of the con- 
denser, C. Cold water is circulated from E to F through the tubes. 
Heat is conducted from the steam through the tube walls to the circu- 
lating water. Sufficient heat is thus abstracted from the exhaust steam 
(about 950 B.t.u. per lb.) so that the steam changes to the liquid state 
and becomes water. The change from steam vapor to liquid water is 
accompanied by a great decrease in volume (about 20,000 to 1, at an 
absolute pressure of 2 in. of mercury) and a corresponding reduction 
in pressure. 

Note. — Surface Condensers Are Generally Used With Steam 
Turbines. A surface-condenser installation is shown diagrammatically 
in Fig. 181. Jet condensers (Figs. 182 and 183), in which the water comes 



wwMi^4w^ ': 

Fig, 181. — Arrangement of equipment in a turbine-driven plant showing surface con- 
denser and auxiliaries. 


Overflow' v i] . ^ uuu..^ i p f f n 

Cold Y^aferSucIfoiv^ / Entminer Circulafinq Pump Varui'im Pum, 
Hof Water 5ucf ion ^ vacuum ^um, 


Fig. 182. — Steam-turbine installation with barometric-jet condenser, C, and cooling 
tower, T. (Worthington Pump and Machinery Corp.) 



in direct contact with the steam, may also be used. The surface con- 
denser is better adapted to maintaining a high vacuum than is the jet 
condenser; also, the surface condenser recovers the feed water in pure 
form. Therefore, in most cases, the surface condenser is the more econ- 

, 'Turbine Generator. 

-"!:'-: • .-'^V^ ■'^: •■^•-:''-^ .■^^^":V'^^-:"^i;"y->^^-:) 

Condenser- - 

-Wafer Discharge From 
Pump To Condenser 

'Centrifugal Circulating- 
Water Pump 


■ '';iZw/$m:mm§ r 

Overf lory Pipe-'' \ -...'• .'■ '•.".''., 

Fig. 183. — Arrangement of a steam turbine, T, with a jet condenser, C . (Schutte & 
Koerting Co., "Multi-jet" condenser with which no air pump or condensate pump is 

omical for turbine service. For economic comparison between the two 
types and also for their construction, care, and operation, see the author's 
Steam Power Plant Auxiliaries And Accessories. 

173. To Compute The Necessary Condenser Surface And 
Cooling-water Requirements For A Steam Turbine, the chart 

of Fig. 184 may be useful. To use the chart, however, certain 
assumptions must be made and certain desirable values must 
be known, as explained below. 

Explanation Of Use Of Chart Of Fig. 184. — The average tempera- 
ture of the cooling-water supply should be first found, either by experi- 


ment, from the weather bureau, or by assumption. This determines, to 
some extent, what vacuum can be profitably maintained. The tempera- 
ture of saturated steam at the absolute exhaust pressure must be from 
25 to 50° F. higher than the cold circulating water — the lower value for 
high-vacuum work (low absolute pressure, say about 2 in. of mercury) 
and the higher value for low-vacuum work, say about 4 in. abs. exhaust 
pressure. In the chart of Fig. 184 the temperatures of exhaust steam 
are plotted with the absolute pressures along the horizontal axis. The 
temperature rise of the circulating water should next be computed. The 
water should not be heated in the condenser to within less than 10°F. 
of the exhaust-steam temperature. The rate of heat transfer should 
next be assumed. This may be assumed at 300 B.t.u. per sq. ft. per hr. 
per degree difference for 4 in. absolute pressure and 350-400 for 2-in. 
absolute pressure. The use of the chart is illustrated in the following 

Example. — Assume that it is desired to condense 10,000 lb. of steam 
per hr. at 2 in. of mercury absolute pressure. Water is available at 
70° F. Since steam at 2-in. absolute pressure has a temperature (Fig. 184) 
of 101° F., the cold circulating water will be 101 - 70 = 31° F. colder 
than the steam. This (see above) is allowable for a 2-in. pressure. The 
circulating water may be heated to 101 - 10 = 91° F. Hence, a 20° F. 
rise in the temperature of the water is permissible. The tubes of the 
condenser are assumed to transmit 350 B.t.u. per sq. ft. per hr, per degree 
difference in temperature. What is the necessary capacity of the con- 
denser in square feet? How much water will be required? 

Solution. — Find the point A (Fig. 184) corresponding to the desired 
pressure and trace vertically to the 70° F. line at B. Then trace hori- 
zontally to the 20 degree rise line at C. The quantity of water for this 
rise is 95 gal. per min. for each 1,000 lb. of steam (as read on the diagonal 
20° line) or 950 gal. per 7nin. total for 10,000 lb. of steam per hr. Now 
trace vertically to the 350 B.t.u. line at D, and thence horizontally to the 
curve at E. The capacity of the condenser may now be read at F. The 
size of the condenser is 127 sq. ft. for each 1,000 lb. of steam per hr. or 
1,270 sq. ft. for the 10,000 lb. of steam per hr. of this example. 

174. In Installing A Condenser To Serve A Turbine, it is 
customary to locate the condenser below the turbine as shown 
in Figs. 181 and 183. A short connection between the turbine 
and condenser serves to minimize the pressure drop between 
the two and also minimizes the possibility of air leaks. Where 
space limitations demand it, however, the condenser may be 
placed on the same floor with the turbine. Figure 182 shows 
a desirable arrangement of apparatus where a barometric jet 
condenser, C, is used with a steam turbine, E, and is supplied 
Dy water which is recooled in a tower, T. All turbine installa- 



Square Feet Surface Per 1000 Lb. 5 + cam 
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4" 3'.5" 3" 2.5" 2" 15" 

Temperatures In Deg. Fahr.,YVi+h Corresponding 
Absolute Pressures In Inches Of Mercury 

Fig. 184. — Graph, based on a steam rate of 1,000 lb. per hr., for determining the 
necessary condenser cooling surface and cooling-water volume required under various 
conditions. (Worthington Pump and Machinery Corp.) 


Lever for 


Valve By 




To Atmosphere 

Fig. 185. — Schutte automatic free ex- 
haust (atmospheric relief) valve. Valve 
disc, C, is raised when the pressure in A 
exceeds the pressure of the atmosphere. 
This pressure is transmitted through the 
small hole, B, in the damping piston, D, 
to the bottom side of valve disc, C, which 
it raises. 

Vacuum Breaker, 
Operated By Float, 
In Case Water Rhes 
In Condensing 


Water And 
Air Discharge 

Fig. 186. — Sectional view of the con- 
denser of Fig. 183 showing the vacuum 
breaker at B. If, when the turbine and 
pump are stopped the water should rise 
into the condensing chamber. A, then 
float C will be thereby raised. This will 
open the valve B which will permit air to 
flow through D into A, 

H-Expansion Joint Comple+e 
Fia. 187. — Copper expansion joint for low-pressure service- 


K 6^~ M 




I-Vertica\ Section Showing 
General Assembly 


Fig. 188. — Westinghouse rubber expansion joint. The sheet-metal baflBe, <S, pro- 
, vides a smooth passageway for the steam. The rubber member, R, is provided with 
the middle support shown in II. Thus, the stresses in R, due to the pressure of the 
atmosphere on the outside of the joint, are small. Member R, can be replaced without 
disturbing any piping or equipment. The spaces, A and B, between R and <S are so 
arranged that they may be kept full of water and so protect the rubber against the high- 
temperature steam whenever the turbine is exhausting against atmospheric pressure — 
as when starting. Connections are also provided for admitting make-up water to these 


tions should be piped with an atmospheric reHef — for free 
exhaust — valve, (Figs. 182 and 185) in the exhaust line. 
This value is provided so that, should the condenser fail to 
function, the turbine may exhaust to the atmosphere. All 
low-level jet condensers should be fitted with a vacuum breaker 
{B, Fig. 186) to prevent the possibility of water being sucked 
up into the turbine at any time. 

Note. — The Methods Of Connecting Condensers To Turbines 
are two: (1) With expansion joints; it is customary for units smaller than 
10,000 kw., and sometimes for larger units, to take care of the upward 
and downward movement of the condenser by using a flexible expansion 
joint between it and the turbine. Copper joints (Fig. 187) have been 
widely used in the past but, due to their short life, they are being replaced 
by telescoping steel or by rubber joints; see the author's Machinery 
Foundations and Erection. Fig. 188 shows a rubber expansion joint. 
(2) Direct connections between turbine flange and condenser flange, or 
with a short nozzle between, are often used on the larger units; the con- 
denser is then mounted on springs so designed that the maximum limits 
of operating conditions — ^that is, high vacuum and non-condensing — ^will 
not cause a strain on the turbine casing flange which is in excess of the 
value specified by the turbine manufacturer. Condenser supports are 
described in the author's Machinery Foundations and Erection. 


1. What is the function of a regenerator as used with steam turbines? What plant 
conditions usually call for a regenerator? 

2. Draw a sketch to show the construction and operation of a Rateau regenerator. 
Explain its operation. 

3. List the piping accessories with which a regenerator should be equipped and give 
the reason for each. 

4. State briefly under what conditions a regenerator is practical. 

5. Describe how boiler-pressure variations may be utilized with a regenerator. 
Draw sketches to show two methods of utilizing the boiler blowdown in a regenerator. 

6. What operating pressure is usually employed in a regenerator? How much above 
and below this pressure should the pressure be permitted to vary? What is the objec- 
tion to employing a slight vacuum in the regenerator? 

7. Describe the process of equalizing the steam requirements of non-condensing 
engine and low-pressure turbine. 

8. How may the necessary weight of water in a regenerator be computed? State 
the formulas. 

9. What is the purpose of employing a condenser in connection with a turbine? 
How does the condenser accomplish this purpose? 

10. What type of condenser is most generally employed with steam turbines? Why? 

11. Explain the process of determining the cooling surface and circulating water 
requirements for a condenser. What values are considered satisfactory for the tem- 
perature difference between the exhaust steam and cold water? Exhaust steam and 
hot water? For the rate of heat transfer? 

12. What are the customary methods of connecting turbines to their condeuser.s? 



1. What weight of water should be stored in a regenerator which is to serve a 1,500-hp. 
low-pressure turbine which uses 25 lb. of steam per hp.-hr. if the regenerator pressure 
may vary between 22 and 25 lb. per sq. in. abs.? The steam supply may be cut off 
from the regenerator for 3 min. or there may be a momentary supply of 2,000 lb. of 

2. If the turbine of Prob. 1 is situated where a liberal supply of cold water is available 
at an average temperature of 60° F., will it be feasible to operate it at an absolute exhaust 
pressure of 1.5 in. of mercury column and, if so, what condenser surface and how much 
circulating water will be required? 



175. The Extensive Use Of The Steam Turbine In Modem 
Industry Is Due Partly To Its AdaptibiUty To All Steam Con- 
ditions. (See Table 29 for classification of turbines according 
to steam conditions.) The relations of the different kinds 
of turbines to the power-plant steam pressures is shown graph- 
ically in Fig. 189. Steam turbines are used not only for the 

[■'High- Pre^eure steam LineFrom Boiler; ' ] 



A) -1 

f V^ 







5teoim.Line: ■•• ■ ■..••.•.1 



B? Ill 


I;'.' Vacuum Line To Condenser ■:/■••';■ ■•'./.| 

Fia. 189. — Showing the relation of turbines designed for various steam conditions to 
the various steam pressures which are used in power plants. 

high-pressure condensing and non-condensing services for 
which steam engines are applicable, but they are also apphed 
to a number of '' special" services — such as bleeder, mixed- 
pressure and low-pressure services — as will be later explained 
(see definitions of the ''special" turbines in Sees. 35 to 37). 
The value of the turbine for these special services is due largely 
to the fact that it derives considerably more power from low- 
pressure steam in condensing service than does a steam 

Note. — When Both Heat And Power Are Supplied By The Same 
Power Plant, it is economical to generate the steam (which will be 


Sec. 176] 



required for heating) at high pressure and run it through a relatively- 
inefficient non-condensing engine or turbine before it is delivered to the 
heating system. When this is done the power thus secured from the non- 
condensing engine or turbine is a sort of byproduct; and only a small 
amount of fuel is burned, in addition to that which would be required for 
heating alone, for its production. On the other hand, if more power is 
required than can be thus obtained, this additional power can, in most 
instances, be most economically obtained with a condensing steam 
turbine. The "special" turbine is particularly useful in improving the 
combined economy of a heating and power plant. 

176. Table Showing How The Requirements Of Any Given 
Set Of Steam Conditions May Be Fulfilled By A Turbine Unit. 


Exhaust steam avail- 

Exhaust steam 

Turbine used 



None • 

Condensing turbine 


Always more than 
turbine needs 


Low-pressure turbine 


Sometimes more — 
sometimes less than 
turbine needs 


Mixed-pressure tur- 
bine or low-pressure 
turbine and regen- 



Always more than 
turbine will supply 

Non-condensing tur- 
bine and reducing 



Sometimes more — 
sometimes less than 
turbine will supply 

Bleeder turbine and 
reducing valve 

Note. — The Requihements In The Above Table Are Assumed To 
Be Normal — Not Emergency Requirements. In case 3, if the 
turbine load is only occasionally too great for the steam supply, it may 
be advisable to "bleed" the live steam line (use a reducing valve) and 
use a low-presure turbine rather than to install a mixed-pressure turbine. 
In case 5, if the surplus of exhaust is only occasional, it may be more 
economical to use a non-condensing turbine than to employ a condenser 
and a bleeder turbine. Where it is indicated above that no exhaust 
steam is needed, it is meant that there are auxiliaries or other 


equipment in the plant which will supply what is needed — none is needed 
from the turbine under consideration. 

Note. — A Condenser Is Always Necessary For ''Condensing," 
*'Low-PRESsuRE," AND "Bleeder Turbines." See Fig. 189. The 
operation of turbines of these types without a condenser would be an 
impossibility. Condensers are generally used also with mixed-pressure 

177. The Relative Amounts Of Heat Energy Which Are 
Theoretically Available For Turbines Operating Under 

1254 Mu. 175 Lb. Per Sq. In. Abe.-IOOt Superheat. 

906 B.tu 

Fig. 190. — Amounts of heat available (given up by adiabatic expansion) from each 
pound of steam for conversion into work by turbines of various types operating under 
typical steam conditions. (It is assumed that, in the mixed-pressure turbine, E, H lb. 
of steam is admitted at boiler pressure; the other H lb. is run through a non-condensing 
engine and admitted to the turbine at 20 lb. per sq. in. absolute. In the bleeder turbine, 
F, H lb. of steam is bled at 20 lb. per sq. in. absolute.) 

Different Steam Conditions may be understood by a study of 
Fig. 190; see also Fig. 261. These values of the heat available 
hold only for the steam conditions indicated, but these con- 
ditions are typical. The amount of heat which is actually 
converted into work is about one-half to three-fourths (depend- 
ing on the sixe of the unit; see Fig. 20) of the values given in 
Fig. 190. It is assumed in Fig. 190 that the low-pressure 
turbine operates on the exhaust from the high-pressure 
turbine. Therefore the low-pressure turbine does not receive 

Sec. 178] 



a full pound of dry steam for each pound of steam admitted 
to the high-pressure turbine. If dry steam is used by the low- 
pressure turbine, the available heat at this vacuum is 171 
B.t.u. per lb. but, if the moisture were removed from the 
steam by using a separator, practically nothing would be 
gained or lost. In actual practice a large condensing turbine 
develops about twice as much power from the same steam as 
does a non-condensing turbine, or as much as does a combina- 
tion of a high- and low-pressure turbine together, in which the 
high-pressure turbine exhausts into the low-pressure turbine. 
For methods of calculating the available energy, efficiency, 
etc. under different steam conditions, see Sees. 10 and 13. 

178. A High-pressure Non-condensing Turbine Is Especi- 
ally Useful under the following conditions, see Sec. 34 for 
definition: (1) When used in conjuriction with a low-pressure 
or exhaust-steam turbine as part of a compound unit. (2) 
When there is usually a demand for all the exhaust steam which is 
produced by the turbine in driving its load. (3) When lack of 
space, water, or other considerations render condensing operation 
infeasible. Non-condensing turbines find extensive applica- 
tion for auxiliary drives (A and B Fig. 206) and small power 
purposes where the steam consumption is of minor importance 
or where the exhaust may be used 
for feed- water heating. The 
non-condensing turbine is sel- 
dom, except in small capacities, 
used alone as a pnme mover 
because it develops only about 
one-half of the power which a 
condensing turbine will develop 
on the same amount of steam. 

179. Turbines Of The Simpler 
Types Are Usually Used For 
Non-condensing Service Where 
All Of The Exhaust Steam Is 
Useful For Heating (Fig. 191). 
Under these conditions, the steam consumption is of com- 
paratively little importance. Velocity-and-pressure-staged tur- 
bines (Sec. 61) having one or two pressure stages are widely 

Atmospheric, , ,,. , „ 
\<= -Exhaust 1-'^^ High-Pressure 
Ste am 5upp/^---\ 

■To Low-Pressure Steam Load 
y^-- Separator 

Fig. 191. — A high-pressure non-con- 
densing turbine, T, piped for service 
where there is demand for more low- 
pressure steam, S, than is suppUed by 
the turbine. 


used for this sort of service. Bucket-wheel and impulse 
re-entry turbines of the axial-flow type are also widely used. 
Turbines of these types are relatively inexpensive in propor- 
tion to the power which they develop but have relatively 
high water rates; see Div. 14. 

180. High-pressure Condensing Turbines Are Useful 
Whenever A Single Unit Is Desired Solely For The Develop- 
ment Of Power. — They (Fig. 192) are built in sizes up to about 


r 1-5 I d e E lev a'+ i on II-5 e c"t i 6 n " A-A." 

Fig. 192. — A small high-pressure condensing turbine piped for service. Usually, 
the most desirable location for the condenser is directly under the turbine rather than 
at some distant location, which is indicated by the above illustration. 

35,000 kw. as single-cylinder units and up to 70,000 kw. as 
compound units. The condensing turbine has become the 
accepted prime mover for all large modern electric generating 
and low-head pumping stations where steam power is used. 
The reason for this is the high efficiency and large power output 
of these turbines in proportion to their size and cost. See Div. 
14 for economies. 

Note. — The Construction Of Condensing Turbines varies greatly 
with the conditions. Single-stage impulse turbines of the single entry 
and re-entry types are sometimes operated condensing. Large con- 
densing turbines for central stations are multi-stage turbines of impulse, 
reaction, or impulse-and-reaction types. 

181. A Bleeder Turbine (Sec. 37 and Figs. 193 and 194) 
may be considered as a high-pressure turbine which can 

Sec. 181 



operate: (1) Condensing, (2) non-condensing, or (3) partly 
condensing and partly non-condensing at the same time. Under 

bleeder Valve 

' Oovernon 

Pressure i.\y 

Exhaust To Condenser 
Througt) Base 


Fig. 193. — Westinghouse automatic bleeder turbine — single-flow type. A vertical 
section of a similar turbine is shown in Fig. 194. 

To Condenser- 

lEnd VievY ''' II-LonojitucJin«l Section 

Fig. 194. — Vertical section and end elevation of a 1,500-kw. W estingliousehX&e^ev 


some conditions it will operate almost wholly as a condensing 
unit; under others, almost wholly as a non-condensing unit. 


The turbine is so automatically controlled that it will: (1) 
Utilize and exhaust into the heating system all steam which is 
admitted by its governor and which is required in the heating 
system; if enough steam for the heating-system requirements 
is not admitted by the turbine governor, high-pressure steam 
may be automatically bypassed into the heating system 
through a reducing valve, R Fig. 191. (2) Utilize and con- 
dense all of the steam which is admitted by its governor in 
excess of that required by the heating system. 

Explanation. — Consider that the bleeder turbine {T, Fig. 195) is 
installed in a plant which requires power all the year for lights and small 

Mtposphenc Relief ^alye-- 

To Condenser. -■'' To Heating System- - '■• 


Fig. 195. — A bleeder turbine installed to supply a low-pressure main with steam and 
condense the exhaust which is not needed for heating. 

motors and requires an amount of heat which varies greatly with the 
changes in the weather. The bleeder turbine is supplied with live steam 
at A. Low-pressure steam for heating is withdrawn at B. The steam, 
in passing from A to B in the turbine, does work which is useful in gen- 
erating power. The steam which is not needed for heating passes on 
through C to the condenser, thus doing more work. In this way the 
heating and power requirements of the plant are satisfied and all of the 
steam is used as economically as is reasonably possible. 

182. The Governing Of A Bleeder Turbine And The 
Proper Distribution Of Steam In It require a regular speed 
governor and a bleeder valve. The turbine and governor 
(see Div. 6) are very similar to an ordinary condensing turbine 
and governor. A bleeder valve (7, Fig. 194 and Fig. 196) 

Sec. 182] 





Sec. 182] 



must however be provided in order to bleed or extract suffici- 
ent steam to maintain a desired pressure in the heating system. 

Explanation. — In the Westinghouse turbine shown in Fig. 194, the 
steam is admitted through governor valve, A, and flows through impulse 
blading, B, and high-pressure reaction blading, Bu Then, if the pressure 
is low in the low-pressure line connected at 0, low-pressure steam is 

Fig. 1965. — Bleeder diaphragm of the Terry turbine which completely stops the steam 
flow through the turbine, diverting it to the bleeder line. Steam returned through the 
bleeder valve (Fig. 196A) enters the nozzles in the upper half of this diaphragm and 
then passes on through the turbine. 

withdrawn through that passage. If the steam pressure increases in 0, 
the valve, V, which is similar in its action to a weight-loaded safety valve, 
opens and allows low-pressure steam to flow through the low-pressure 
blading, 5 2, to the condenser. A check or non-return valve is always 
provided in the low-pressure steam line to prevent flow of steam back to 
the turbine. 

Note. — The Genebal Electric Co. Bleeder Mechanism is shown 
in Fig. 196. The bleeder or extraction valve consists of a diaphragm, 
D, placed across the turbine cylinder at the point where it is desired to 
bleed the turbine, and a valve disc, V. The diaphragm and disc are so 


arranged that, as the disc is rotated part of a revolution, the slots, S, 
in the diaphragm (through which steam is admitted to various nozzle 
sections) are uncovered successively. That is, a slight rotation will 
uncover one slot; a larger rotation will uncover two, three, or all of the 
slots. The rotation of the disc is controlled by the piston and relay 
mechanism, PRX. Steam from the low-pressure line is admitted behind 
the spring-opposed diaphragm, X. The movement of this diaphragm 
operates the piston, P, through the oil-relay valve, R. In this way the 
opening of the relay valve is controlled by the steam pressure in the 
low-pressure line. By adjusting spring, L, this pressure may be main- 
tained at any reasonable desired value. The advantage claimed for 
this method of extraction over that of Fig. 194 is that with the Fig. 196 

-Low-PresSure Reacf'on Blading-. ^^ 
High -Pressure Impulse Blading- 

Fig. 197. — Vertical section of a mixed-pressure tu 

method there is little throttling action in the bleeder valve since most of 
the slots, S^ are always either wide open or tightly closed. 

Note. — The Bleeder Mechanism Of The Terry Turbine is 
shown in Fig. 196A. It differs from the bleeder mechanisms just de- 
scribed in that the steam flow through the turbine is completely stopped 
off by a special diaphragm, Fig. 196B. The steam, after flowing through 
the first stages of the turbine, is diverted by this diaphragm (D, Fig. 
196A) into the low-pressure steam pipe, L. Should the pressure in this 
pipe become too great, it will displace a diaphragm in the regulator, R, 
and thereby open an oil-relay valve. Oil will then flow through the 
relay valve to a piston on the same rod as the bleeder valve, F, thus 
opening Y, Steam will then flow through F, again into the turbine — 
now through the low pressure stages. Should the pressure in L become 
too low, the reverse action takes place — valve Y is closed. The valve, 
F, is so proportioned that, should its operating mechanism become in- 

Sec. 183] 



active, it will automatically open at a predetermined pressm-e in L, thus 
avoiding dangers due to excess pressure. 

183. A Mixed-pressure Turbine (Sec. 36 and Fig. 197) may 
be considered as a combination, in a single machine, of a high- 
pressure and a low-pressure condensing turbine. A mixed- 
pressure turbine is so controlled that no high-pressure steam 
will be used unless the low-pressure steam supply is inadequate 
for the power requirements of the turbine at that instant. 

Explanation. — Consider that the mixed-pressure turbine {T, Fig. 
198) is installed to utilize the exhaust steam from the engine, E. Exhaust 

l/ve' High-Pressure 
f/^~^ Sfeam Main \ 

tlixed Pressure 

j^//f yy//^/^ /^// v/^/ y/// /yf/^ //y /w //^^ /y/ //// y/M /^-/\ 

Fig. 198. — Mixed-pressure turbine installed for service in connection with a recipro- 
cating engine. 

steam is admitted to the turbine at A and flows through it to condenser, 
C. If the load on the engine is heavy and that on the turbine is light, 
the turbine runs as a low-pressure turbine, and the surplus exhaust steam 
from the engine is condensed. Now suppose that the load on the engine 
becomes very light and that on the turbine becomes very heavy. The 
turbine will then derive little power from the engine exhaust and would 
stop if no other source of power were available. But the governor of the 
turbine then admits high-pressure steam at B which flow^s through all of 
the stages of the turbine. The turbine will then derive most of its power 
from the high-pressure "live" steam just as does a high-pressure con- 
densing turbine. 

184. The Functions Of A Governor For A Mixed-pressure 
Turbine (see Div. 6) are: (1) To admit all available loio-pressure 


steam provided it is all required by the turbine in delivering its 
load. (2) To shut off the low-pressure steam if more than sufficient 
for the load. (3) To admit just sufficient additional high- 
pressure steam to carry the load when the low-pressure steam 
supply is inadequate. 

Explanation. — These functions may all be accomplished very simply 
by the arrangement shown diagrammatically in Fig. 199. The governor 
is shown in I (Fig. 199) in the no-load position, with the weights or balls 
raised. When load is applied to the turbine and its speed decreases, the 

Governor ■ 
Adjustable Limif 
:0f Tra\/el 


Y' Valve 
L M ^Hl 

1-No Load, Speed Hfgh. 
(Both Hicjh And Low 
Pressure Valves Closed.) 

^ L ^ "mm 

E-Normoil Load, Normal Speed. 
(low Pressure Valve Open To 
I+s Liml+.The High Pressure 
Valve Remaining Closed 
Because Of The Weight) 

Fig. 199.- 


E-Same Load And Speed 
As In lE.(Fai lure Of Low- 
Pressure 5+eoim Closes 
Low-Pressure Valve And 
In Turn Opens The High 
Pressure Valvej 

Diagrammatic representation of the operation of a governor for a mixed- 
pressure turbine. {Terry Turbine Co.) 

m-FulI Load, Speed Low. 

(More Power Is Required Than 
Can Be Supplied By Low- 
Pressure S+eam, Consecien-rly 
High Pressure Valve Open^ 

balls drop, as in II, and lift pivot, P. Due to the weight, W, the move- 
ment of P lifts the low-pressure valve, L, but high-pressure valve, H, 
is held on its seat as in II (Fig. 199). After the low-pressure valve has 
traveled as much as the adjustable stop, S, will permit, as in ///, further 
movement of the governor lifts high-pressure valve, H, against the 
downward force of W. If it is desired to maintain a certain back pressure 
in the low-pressure steam line, an automatic travel regulator, T (Fig. 
199, 77) must be employed. This consists of a cylinder containing a 
spring-loaded piston. If there is no pressure in the lower part of T, no 
travel of the low-pressure valve is permitted and the turbine runs on high- 
pressure steam. But if a pressure is produced in the lower part of T, the 
lifting of L is permitted so that low-pressure steam is admitted to the 

Sec. 184] 



turbine. An actual mixed-pressure governor valve is shown in Fig. 200. 
This arrangement never closes the low-pressure valve when there is load 
on the turbine. If it is desired to maintain a back pressure, a constant- 
pressure valve (Fig. 201) must be used. This valve also acts as a check 
to prevent a flow of steam from the turbine to the low-pressure line in 


' 5econo/arL/ 
\ Link. 

Opera fin^ 
Piston Rod.. 


Fig. 200. — Governing valves of a mixed-pressure turbine. As oil is admitted from 
the relay valve (not shown) to the under side of the piston, P, lever A is rotated upward 
and to the left with the link, C, which is pivoted at B. This raises and opens the low- 
pressure valve, L. At a certain point in the upward motion of P, the lost motion in 
link D is taken up. Further upward motion of P will also open the high-pressure valve, 
H. No provision is made in this governing mechanism for keeping L closed when the 
pressure in the low-pressure steam supply-pipe becomes abnormally low. 

case the low-pressure steam supply fails. See also Fig. 202. 

Note. — ^Low-pressure Steam Is Sometimes Supplied To The 
Later Stages Of An Ordinary Condensing Turbine Through Only 
A Flow Valve (Fig. 201). — Turbines which are so arranged are not 
generally called mixed-pressure turbines although they really function as 
such. The low-pressure steam is admitted by the flow valve whenever 
the pressure in the supply pipe (the exhaust pipe of the non-condensing 


equipment) exceeds a predetermined value. There is no speed-governor 
valve to control the admission of the low-pressure steam. Hence, 
should such a turbine be run under very light load at a time when the 
low-pressure supply is plentiful, the turbine may run at a speed well 


Hand-Wheel For Lowering Or 
Ro/i's/ngr Pressure Plate 


Piston Chambers 
"null ': 


Fig. 201. — Cochrane "constant-pressure" multiport flow valve (reducing valve) 
used at the low-pressure inlet of a low- or mixed-pressure turbine. This valve is used 
to maintain a constant back pressure on a non-condensing unit. This valve may be 
set to maintain the desired constant pressure by turning //, which changes the 'compres- 
sion of the springs, <S. If the pressure in A falls below that for which the valve is set, 
steam pressure in B lowers valve discs, Y , and shuts off steam from B. If the pressure 
in A increases above the pressure for which the valve is set, the pressure in A lifts the 
valves, V , against the springs, &. At Z) is a dashpot which prevents chattering and 
above it is a buffer spring. 

above normal. To prevent such overspeed damage to the turbine, the 
low-pressure supply is shut off by the automatic overspeed governor 
when the turbine's speed reaches the value at which this emergency 
governor is set to operate. Hence, such turbines should be used only where 
there is very little likelyhood that the low-pressure steam supply u^ll ever 
exceed the requirements of the minimum load on the turbine. 

Sec. 185] 



185. Mixed-pressure Turbines Are Sometimes Used For 
Auxiliary Drives. Figure 203 shows mixed-pressure main 
turbine, T, and auxiliary turbine, A, so connected that they 
may derive steam from the receiver, R^ of a compound engine. 
These turbines running condensing are considerably more 



Fig. 202. — Schutte & Koerting automatic flow regulating valve. This valve is, in 
function, similar to that of Fig. 201. This valve will, however, maintain a constant 
pressure on its supply side regardless of the pressure on its discharge side and without 
manual adjustment. On the other hand, this valve does not serve as a check valve 
whereas that of Fig. 201 does. The rubber diaphragm, R, is supported by plate, B, 
and is submerged in water to protect it from the hot steam. Multiplying levers connect 
B with the valve spindle, S. The valve is shown in the closed position, which it normally 
occupies when the pressure above R is less than about 16 lb. per sq. in. abs. A greater 
pressure above R will cause it to lower the valve discs, D, and raise the weights, W, on 
levers, L. Steam may then pass through the valve to the turbine. Should the pressure 
above R fall below 16 lb. per sq. in. gage the valve will be closed by the weights, W. 
The valve may be blocked in the closed position by screwing up wheel A. 

economical than the low-pressure cylinder, L, of the engine. 
Thus for most loads on the engine, auxiliary power is secured 
with a negligible amount of extra steam. When there is an 
overload on the engine or when the engine is not running, live 
steam may be admitted, through M and N, to the main and the 
auxiliary turbines. 

186. There Are ANumber Of Automatic Or Partly-automatic 
Methods Of Balancing The Heat And Power Requirements 
Of A Steam-turbine Power Plant. — In some of these, (Sec. 184) 


the automatic balancing is accomplished by employing valves 
which are sensitive to variations in the pressure of the low- 
pressure steam. In others, the balancing is accomplished by 
electrical or mechanical means. See examples below. 

Note. — "Heat Balance" Or A Balance Between The Power 
Required For Auxiliary Drives And Heat Required For Feed- 
water And Other Heating is an important consideration in most 

Hot-Well Pump-' "-Circulating Pump' 

Fig. 203. — Mixed-pressure turbines, T and A, arranged to operate on steam from the 
receiver, R, of a cross-compound condensing engine. 

plants. From about 2 to 10 per cent, of the steam generated by the 
boiler is generally required, directly or indirectly, to drive the auxiliaries 
of the power plant. About 5 to 8 per cent of the steam generated by the 
boiler may — after it has been used in some non-condensing engine or 
turbine — be profitably used for heating the feed water. Sometimes, 
therefore, if all the auxiliaries are steam driven, they will supply enough 
— or more than enough — exhaust steam for feed-water heating. For 
maximum economy, there should, theoretically, be just enough exhaust 
steam available for feed-water heating but there should be no waste of 
exhaust. The temperature to which the feed water should be heated 

Sec. 186] 



■ ■ • ■ • ■ . ' . ■ ' circulating' Pump- -Hof-Well Pump' 

Fig. 204. — Heat-balance system with bleeder turbine prime mover and back-pressure- 
turbine driven auxiliaries. {De Laval Steam Turbine Co.) The back-pressure turbines, 
B and B, operate on live steam from the boilers. They exhaust into the feed-water 
heater, H, against a back-pressure. The flow valve, V, permits steam to flow from the 
extraction chamber of main turbine, T, into the heating system whenever the difference 
between the pressures in the two exceeds the value for which V is set. Thus as the load 
on T varies, the pressure in the heating system may also vary unless V is adjusted by 
the operator. For periods when the load on T is very small, a reducing valve (not 
shown) may be necessary to admit live steam to S. 

To Atmosphere' 



nnnnF^ rrnr:: 

Hot-Well Pump- 

v/// 'OA 

Fig. 205. — Heat balance system in which two bleeder prime-mover turbines, T, (only 
one is shown) are used and in which the auxiliary drive turbines, A, are of the mixed- 
pressure type. 


with exhaust steam is about 210° F. where no economizer is used. Where 
an economizer is used, the feed-water temperature should vary from 
about 210 to 150° F. as the water flows from the exhaust-steam heater. 
See the author's Steam Power Plant Auxiliaries and Accessories. 

Example 1. — In the arrangement of Fig. 204, both the bleeder turbine, 
T, and the back-pressure-turbine (Sec. 34) driven auxiliaries, B and 
B, are connected to supply steam to the feed-water heater, H, and to the 
steam-heating system, S. When the steam from the auxiliaries is ample 
for all steam-heating requirements, no steam will flow from the bleeder 
turbine, T, to the heating system. All of the steam which the turbine, 
T, then uses will be condensed in C. Thus the bleeder turbine itself 
furnishes an automatic means of keeping the heating requirements and 
the low-pressure steam supply balanced. 

Example 2. — Figure 205 shows an arrangement in which the main tur- 
bine, T, is a bleeder and the auxiliary drives. A, are mixed-pressure tur- 
bines. For very heavy heating loads, the main turbine is run entirely 
non-condensing and exhausts to the heating system, S. When there is no 
heating load, the main and auxiliary turbines are both run entirely con- 
densing. When there is a moderate heating load, steam is bled from the 
main turbine to supply both the heating system and the mixed-pressure 
auxiliary turbine with low-pressure steam. If the power load is increased 
so that it cannot all be handled thus, the auxiliary turbine may run 
entirely on high-pressure steam and exhaust to the heating system. This 
will permit condensing all of the exhaust from the main turbine so that its 
maximum power will be developed. 

187. An Electrical Method Of Effecting An Exhaust-steam 
Heat Balance In A Power Plant is shown in Fig. 206 (from 
Power, Sept. 6, 1921). This method is applicable either for 
plants which are used for developing electrical energy only 
or for combined heating and power plants. 

Explanation. — The main turbine, T, is operated condensing. In 
order that the power-plant lighting and motor drives may not be affected 
by trouble in the main electric system, a non-condensing house turbo- 
alternator, H, is employed to generate the necessary electrical energy 
which is used in the power plant itself. The motors, M, form part of the 
electrical load on H. The exhaust steam from H is piped to the baro- 
metric feed-water heater F. But, since the electrical load on H cannot 
readily be varied, it is obvious that the amount of exhaust steam for 
feed-water heating will be nearly constant unless some variable load is 
connected to H. If H were paralleled with the main generator (by 
connecting it to the main bus bars), then the load on H could be varied 
by varying its governor-spring tension (Div. 6) thus causing it to furnish 
more or less power to the main bus bars But this would place the power- 
plant lighting and motor drives subject to shut down due to trouble on 

Sec. 187] 



the main lines. To obviate this possibility, the motor-generator G is 
introduced as a connecting link between H and the main lines. 

The temperature in the feed tank W is recorded by a remote-reading 
thermometer on the switchboard. The switchboard operator, by 
manipulating the synchronizing motor on H may then cause H to 

C5 -^ 

ft 3 

M O 

ft 03 

IN .2 

deliver power to or the house system to take power from the main bus 
(through G) and thus exhaust more or less steam as required for feed- 
water heating. A definite feed-water temperature, which has been 
found most economical, may thus be maintained. The exhaust steam 
from the non-condensing turbines, A and B, which drive the auxiliaries, 
is used for distilling make-up water. Any exhaust which is not thus 
used flows through the relief valve, R, to the heater, F. 


188. The First Costs Of Mixed -pressure And Bleeder 
Turbines Are Relatively Low compared to those of separate 
equipment for the functions which these turbines perform. 
A bleeder turbine takes the place of a condensing and a non- 
condensing turbine. It also furnishes automatic means of 
conserving steam. A mixed-pressure turbine may take the 
place of an exhaust-steam turbine and a high-pressure turbine. 
Considered in another way, the mixed-pressure turbine devel- 
ops power from exhaust steam and obviates the necessity for a 
regenerator by drawing live steam when the supply of exhaust 
is low. The cost of this live steam may often be neglected 
because the times when it is used are those intervals just after 
the non-condensing equipment has been shut down — at such 
times there is likely to be a surplus of steam and the safety 
valves of the boilers would blow if no steam were drawn from 
the boilers. 

Note. — The Speed Regulation Of Mixed-pressuke Turbines And 
Bleeder Turbines (see note under Sec. 125 for definition of "speed 
regulation") is ordinarily much greater than that of other turbines. 
In bleeder turbines, the governor valve must open somewhat wider than 
in ordinary turbines to admit sufficient steam to develop the full power 
of the unit when the bleeding is heavy. This necessitates more travel 
of the governor and valve and more variation in speed. In mixed- 
pressure turbines, the governor gear must travel far enough to open the 
low-pressure valve and far enough in addition to open the high-pressure 
valve when there is little exhaust steam. This travel requires a greater 
governor movement than would be required to admit steam from a single 
source. Also the speed regulation of mixed-pressure and of bleeder 
turbines is Ukely to be slightly different when considerable low-pressure 
steam is being used or extracted from that when little low-pressure steam 
is being used or extracted. 

189. The Economies Of Bleeder And Mixed -pressure 
Turbines are calculated from two different standpoints: 
(1) A technical standpoint. From a technical standpoint, the 
economies of mixed pressure and bleeder turbines are most 
conveniently calculated on a basis of available heat and 
efficiency ratio as in Sec. 15. The efficiency ratio of these 
turbines and of low-pressure turbines when operating near the 
capacity for which they are designed is about the same as 
that of high-pressure condensing and non-condensing turbines 

Sec. 189] 



of the same capacities. (2) A commercial standpoint. An 
example of how the steam consumption of a bleeder turbine 
may be considered commercially is shown in Fig. 207; the 
turbine is, from this standpoint considered to consume only 
that steam which it condenses. The consumption is con- 
sidered to be the net consumption, or that fed to the turbine 















— . 





Lb. Per 

















) 12 

5 2 



/5 5 

)0 6' 

.5 7 

)0 6 

75 IP 

00 M 

25 U 

50 i,y 

rS 1,50 

Load, Kw. 

Fig. 207. — Graphs showing variation in commercial economy of a 1,000-kw. bleeder 
turbine with variations in load and rate of bleeding. 

at boiler pressure minus that bled from the turbine at a low 
back pressure. For commercial purposes this assumption is 
not much in error because the steam which is bled has 90 
to 95 per cent, as much heat as has the high-pressure steam. 
When the turbine is bled heavily and is carrying a light load, 
its "commercial" steam consumption may, on this basis 
as shown, Z (Fig. 207), be practically zero. Similarly a mixed- 
pressure turbine would in some instances, where there is a 
surplus of exhaust steam from non-condensing equipment, 


be charged with only the live steam which it used. Then its 
steam consumption, from a commercial standpoint, might be 
zero most of the time. 

190. To Compute The Approximate Rate At Which A 
Mixed-pressure Or Bleeder Turbine Consumes High- 
pressure Or Live Steam, use the following formula : 
(30) Wbi = 

^^^}cw ~ '^''^^' ~ ^''^] ^^^- p^'^p- ^'-^ 

Wherein: Wsi = the weight of high-pressure steam, in pounds 
per brake horsepower-hour, which passes through all of the 
stages of the turbine. Wb2 = the weight of low-pressure 
steam which is admitted to a mixed-pressure turbine or 
which is extracted from a bleeder turbine, in pounds per brake 
horsepower-hour. Hi and H2 = respectively, the inijbial 
and final total heats per pound of steam before and after 
adiabatic expansions, of the high-pressure steam, correspond- 
ing to the weight W^i. H/ and H2' = respectively, the initial 
and final heats per pound of steam, corresponding to Wb2. 
Er = the efficiency ratio, or ratio of the actual efficiency to 
that of the ideal Rankine cycle; this is the value plotted in Fig. 
20. The heat values are found on the graphs of Fig. 15, 
exactly as explained in Sec. 15 for regular high-pressure 

Example. — A 2,000 hp. mixed-pressure turbine consumes at full load, 
9 lb. of steam per horsepower-hour at atmospheric pressure. The 
condenser maintains 28.5 in. of vacuum. How much high-pressure steam 
at 175 lb. per sq. in. gage and 100° F. superheat will it also consume at 
full load? Solution.— From Fig. 20, E^ = 0.65. Also, from Fig. 15, 
Hi = 1,256; H2 = 888; H/ = 1,150; Ho' = 965. Hence, by For. (30): 
Wbi = [l/(Hi - H2)][(2,545/E,) - Wb^CHi' - H2')] = [1 - (1,256 - 
888)] X {(2,545 ^ 0.65) - [9 X (1,150 - 965)]} = 6.12 Ih. per hp. hr. 

Example. — A bleeder turbine, which operates on saturated steam at 
165 lb. per sq. in. gage, supplies a heating system which requires 12,000 
lb. of steam per hour at 5 lb. per sq. in. gage. The turbine is rated at 
1,000 hp. The condenser maintains a 29-in. vacuum at full 
load. What will be the total steam consumption of the turbine in pounds 
per hour at full load? Solution. — From Fig. 20, Er = 0.60. From 
Fig. 15, Hi = 1,196; Ho = 835; H'l = 1,196; H2' = 1,034. From the 
given data, Wb2 = 12,000 -r- 1,000 = 12 lb. per hp.-hr. Hence, by 
For. (30): Wbi = [l/(Hi - H2)] [(2,545/E.) - Wb2(Ri' - H2O] = 


[1 -^ (1,196 - 835)] X-i(2,545 ^ 0.60) - [12 X (1,196 - 1,034)]} = 
6.37 lb. per hp. hr. Hence, the total steam consumption of the turbine 
= 6.37 X 1,000 + 12,000 = 18,370 lb. per hr. 

191. To Compute The Steam Consumption Of A Bleeder 
Turbine At Any Load And Any Rate Of Bleeding when its 
consumption at various loads with no bleeding is known, use 
the graphs of Fig. 208 (Joseph Gershberg in Power, Oct. 11, 
1921). It may be safely assumed that the economies of a 
bleeder turbine which is not bled are very nearly the same as 
those of a high-pressure condensing turbine of the same size 
and type. The diagram is limited in its application to turbines 
of 300 to 2,500 kw. capacity using steam at 125 to 150 lb. 
per sq. in. gage, bleeding at to 20 lb. per sq. in. gage and 
condensing at 26 to 283-^ in. of mercury. 

Explanation. — The fraction B/Fiqo, which is laid out on the horizontal 
scale of the diagram, is first calculated. B/Fioo = {the rate of bleeding of 
the steam, in pounds per hour) -=- {the steam consumption of the turbine at 
full load — no bleeding — in pounds per hour). This value is then found 
on the scale and followed vertically until the inclined-line graph is inter- 
sected which corresponds to the percentage of full load at which the 
consumption is to be calculated. The point of intersection is then 
projected and a value of the fraction Fb/Fc is read on the vertical scale. 
Fb/Ec = {the consumption with bleeding at the rate B) -^ {the consumption 
without bleeding at the same load). The consumption without bleeding, 
multiplied by this Fb/Fc ratio, will give the consumption at the given 
rate of bleeding. See the following example. 

Example. — A turbine uses 10,000 lb. of steam per hour at full load and 
6,000 lb. at half load, when there is no bleeding. What will be the 
consumption at H load when bleeding 5,000 lb. per hr.? Solution. — 
Calculate B/Fioo = 5,000 -^ 10,000 = 0.5. Find 0.5 on the horizontal 
scale as indicated by the dotted line and trace up to where the 50-per 
cent.-load graph is intersected at A. Then move to the left and read the 
value of Fb/Fc, which is found to be 1.56. The consumption at half 
load with this rate of bleeding is then 6,000 X 1.56 = 9,360 lb. per hr. 

192. Exhaust-steam Or Low-pressure Turbines Are Appli- 
cable under several conditions (see Sec. 35 for definition): 
(1) To improve the economies of a condensing reciprocating- 
engine plant. (2) To utilize the exhaust steam from non-con- 
densing reciprocating machinery. (3) As part of a compound 
unit, to run from the exhaust of a high-pressure turbine. 






























































































\ ^ 

^ / 





































































Fig. 208. — Diagram for estimating the steam consumption of bleeder turbines. 
For turbines 300 to 2,500 kw., 125 to 150 lb. steam pressure, 26 to 28>2 in. vacuum, 
steam bled from to 20 lb. gage pressure. 

B = amount of steam bled in pounds per hour at any load. 
Fioo = amount of high-pressure steam required in pounds per hour at full load when 
no steam is bled. 
Fb = Total amount of high-pressure steam required in pounds per hour when bleed- 
ing B pounds per hour at a specific load. 
Fq = Total amount of high-pressure steam required in pounds per hour for the same 
specific load when no steam is bled. 

Sec. 193] 



Exhaust-steam turbines are usually either of the double-flow 
reaction (Fig. 79) or the Rateau type, although the single-flow 
reaction type (Fig. 209) is also used. The double-flow feature 
is used in the reaction type because of the large volume of 
steam which must be accommodated at the low pressure. 
The large volume is accommodated in turbines of the Rateau 
type by making the nozzle area proportionally large. 

Note. — ^Low-pressure Turbines Always Operate Condensing. — 
There is so little power available between the usual pressure of low- 
pressure steam (0 to 15 lb. per sq. in. gage) and atmospheric pressure that 

BcilancQ Pisfon i^^ar/n^ 

Fig. 209. — Allis-Chalmers single-flow low-pressure turbine. 

no turbine would be justified for low-pressure non-condensing service. 
Note. — Irregular Supplies Of Exhaust Steam Cannot Be 
Utilized Satisfactorily By A Low-pressure Turbine Alone. — When 
the supply of exhaust steam on which the turbine is to operate is irregular 
— as when the source is a steam hammer or a rolling mill engine — some 
means, such as a regenerator, of storing or accumulating a supply of this 
steam is sometimes used, (see Div. 8). Another method is to employ a 
mixed-pressure turbine; then the deficiency in exhaust steam is made up 
by drawing live steam from the high-pressure steam line. 

193. The Addition Of A Low-pressure Turbine Usually 
Improves Both The Capacity And Economy Of An Existing 
Non-condensing Reciprocating-engine Installation (Fig. 210). 
The increase in capacity is usually 75 to 100 per cent. That is, 


if the non-condensing engines develop 1,000 hp., the engine- 
and-turbine combination may develop 2,000 hp. The increase 
in economy, expressed as a percentage of the water rate is 
usually 30 to 50 per cent. That is if the engine operating 
non-condensing consumes 30-lb. steam per brake horsepower- 
hour, the engine-and-turbine combination may consume only 
15 lb. per b.hp.-hr The first cost of a low-pressure turbine 

To Atmosphere-, 
Mulfiporf -^ 

■Relief Yalre 

Pump Ex hoiu$y '' '^ Sfeani Trap''' ' '^Feed-Wafer Heater Anol' Receirer" ". ■"■ 

Fig. 210. — Low-pressure turbine, T, installed to operate on the exhaust from a non- 
condensing reciprocating engine, E. 

is very low compared to the cost of an additional boiler and 
high-pressure unit for the same amount of additional power. 

Note. — It Is Generally Well To So Arrange That Each Engine 
Will Supply Its Own Separate Low-pressure Turbine And Con- 
denser, principally because, if one turbine and condenser served several 
engines, condenser or turbine trouble would render the entire outfit 
ineffective. Where there are a number of very small units, it may be 
better to provide but one turbine-and-condenser for a group of two or 
three engines to insure minimum first cost per kilowatt capacity. In 
any case, there should, preferably, be more than one complete low-pres- 
sure-turbine-condenser unit in each plant so that the danger of a complete 
breakdown will be a minimum. If several engines exhaust to one turbine- 
and-condenser, each engine should always in starting be run non-con- 
densing a few strokes. This is to avoid impairing the condenser vacuum 
with the air which was in the engine cylinder when it was lying idle. 

Sec. 194] 



Note. — Receivers And Steam-and-oil Separators Should Ordi- 
narily Be Installed Between Engines And Mixed- Or Low- 
pressure Turbines; see S, Fig. 210. The water and oil which is present 
in the engine exhaust may do comparatively little damage to the turbine 
if the oil is pure— except that they increase the friction of the turbine 
blading. But if the oil is impure and contaminated with matter taken 
mechanically from the boilers, it may form deposits on the turbine blades 
and thus seriously interfere with the operation of the turbine. A 
receiver is usually necessary to equalize the pulsations in the steam supply 
which result from the intermittent exhaust from the engine. In Fig. 210, 
the open feed-water heater, W, acts as a receiver. 

Three Phase- - 
Current Bus 


Rec iproca fing 


Alfernafirig Current 

'Zxhaust From Engine 

Fig. 211. — Diagram showing method of operating reciprocating-engine and low-pres- 
sure-turbine generating units on the same alternating-current line without, governing 
the turbine. 

194. Several Methods Of Balancing The Load Between A 
Non-condensing Reciprocating Engine And A Low-pressure 
Turbine are shown in Figs. 211, 212, 213 and 214. It is 
desirable to have the engine in such installations produce 
exactly as much exhaust steam as the turbine requires. Then, 
all of the steam will be used with maximum economy. 

Example 1. — When (Fig. 211) both the low-pressure turbine, T", and 
the non-condensing engine, E, drive alternating-current generators, Gi 


and (72, which are both connected to the same alternating-current Hne, 
the arrangement is self-balancing. The two units are automatically 
by electrical interaction between the generators, kept at the same 
synchronous speed. If their load characteristics are similar, the steam 
from the engine will always be just sufficient for the turbine — when once 
the loads have been balanced. There will not be any excessive variation 
of the exhaust steam pressure in receiver, R. The turbine may then be 
run without a speed governor. The engine governor and the turbine 
emergency governor serve to control the speed and provide protection. 
Example 2. — When (Fig. 212) the engine drives a direct-current 

Fig. 212. — Method of supplying both alternating and direct current from turbine 
and reciprocating-engine generators, using synchronous converter for balancing the 

generator and the turbine an alternating-current generator or vice versa, 
the conditions are essentially the same as in Example 1 above except that 
a synchronous converter, S, must be employed to balance the alternating- 
and direct-current loads. There will be voltage fluctuations when the 
converter changes its function from maintaining the alternating-current 
voltage at the expense of the direct-current to the reverse operation but 
this fluctuation may be corrected at the switchboard. 

Example 3. — Where (Fig. 213) the mechanical load, Lm, on the 
engine, E, and the electrical load, Ljj, on the turbine, T, are balanced by 
means of a synchronous motor, M, the two units may be controlled by one 
governor as in Examples 1 and 2. Some adjustment at the switchboard 
is necessary when the motor changes over from acting as a motor to 
acting as a generator. 

Non-Conofensi'ngr Engine .. _ 


Flywheel- - . ^ 

Line 5 haft- ^ 






Current L'me5\ 


Fig. 213. — Showing how mechanical and electrical loads may be interconnected so 
that the power requirements of a mechanical-drive engine and the low-pressure turbo- 
generator which it supplies with steam will be balanced. 

' Pisfon 



Fig. 214. — Transverse section 
showing the governing valve of a 
low-pressure turbine. Exhaust 
steam is admitted through valve 
X which is controlled as the oper- 
ating piston P is actuated by oil 
from the governor relay valve. 
When X is wide open, further 
movement of P admits high-pres- 
sure steam through the valve V. 


Example 4. — Where (Fig. 215) the two loads, Li and L2, are entirely- 
separate, both units, E and T, must be governed independently. The 
engine, E, will then furnish much more steam at times than the turbine, 
T, requires. The excess is automatically passed through the flow valve, 

Three-phase A.C. Une—--^ L2 

-Loyr- Pressure Turbine 


'■^Separating Receiver 
^"Non-Condensing Engine 




Fig. 215. — Application of a low-pressure turbine where a reciprocating engine drives 
a line shaft and always furnishes enough exhaust steam for the power requirements of 
the low-pressure turbine. The excess steam from the engine which is not needed by 
the turbine is condensed. 

Y (see Fig. 201), and is condensed. If the engine exhaust is occasionally 
insufficient for the turbine, a live-steam valve (F, Fig. 214) on the tur- 
bine will open and permit the deficiency to be made up; the low-pressure 
turbine then performs the function of a mixed-pressure turbine in a way 
but has, of course, no high-pressure blading. 


1. Name three special applications of steam turbines in power plants for which 
steam engines cannot be economically used. Two for which engines can also be used. 

2. Why is a non-condensing turbine useful when much low-pressure steam is needed 
for heating? 


3. Where no exhaust steam is available and none needed, what kind of turbine is 
ordinarily used? 

4. If a condensing steam turbine develops 2,000 kw. on a given supply of steam, 
approximately how much power would a non-condensing turbine develop from the same 
steam supply under typical conditions? 

5. Name two applications of a high-pressure non-condensing turbine. What types 
of turbines are preferred for each application? 

6. A bleeder turbine combines the functions of what two other kinds of turbines? 

7. What two devices are necessary for the governing of a bleeder turbine? What 
function must these two devices perform besides that of keeping the turbine speed 

8. A mixed-pressure turbine combines the functions of what two other kinds of 

9. What are the functions of a mixed-pressure turbine governor? 

10. What is the purpose of an automatic travel regulator for a mixed-pressure turbine 

11. What is the purpose of maintaining an automatic exhaust-steam heat balance in 
a power plant? 

12. Show by a sketch how a mixed-pressure turbine may be connected to other power- 
plant equipment for maintaining an exhaust-steam heat balance. 

13. Show by a sketch how a bleeder turbine may be connected to other equipment for 
maintaining an automatic exhaust-steam heat balance. 

14. How do the costs of mixed pressure and bleeder turbines ordinarily compare with 
those of the other equipment which they can replace? 

15. How do the speed regulations of mixed pressure and bleeder turbines ordinarily 
compare with those of other turbines? 

16. On what two bases are the economies of mixed pressure and bleeder turbines 
considered? Explain how their steam consumptions may sometimes be practically zero 
on one basis. 

17. When is a low-pressure turbine useful? Why is it sometimes economical to install 
one in a condensing reciprocating engine plant? 

18. How much improvement in economy and capacity may usually be expected from 
the installation of a low-pressure turbine in a non-condensing reciprocating-engine plant? 

19. What is the disadvantage of having all the engines in a plant exhaust to one low- 
pressure turbine and condenser? 

20. Why are a steam separator and receiver advisable between an engine and a low- 
pressure turbine? 

21. Show by a sketch how a low-pressure turbo-alternator is connected for parallel 
operation with an engine-driven alternator. 

22. How may the load be balanced between an engine-driven direct-current generator 
and a low-pressure turbo-alternator? Explain with a sketch. 

23. How may the load on a low-pressure turbo-alternator be balanced with that of 
an engine which is used for a line-shaft drive? 


1. In a power plant where the boilers deliver stean, dt 150 lb. per sq. in. gage and 50° F. 
superheat, the non-condensing steam engines consume 6,000 lb. of steam per hour and 
exhaust at a back pressure of 5 lb. per sq. in. gage. It is desired to utilize this exhaust 
steam in a 500-hp. mixed-pressure turbine which will exhaust into a vacuum of 28.5 in. 
of mercury column. About how much high-pressure steam will this turbine require 
per hour when operating at full load? 

2. A 1, 500-hp. bleeder turbine is to take steam at 180 lb. per sq. in. gage and 100° F. 
superheat. It will exhaust into a surface condenser where the vacuum will be main- 
tained at 29 in. of mercury when the barometer stands at 30 in. It will also be required 
to supply 22,500 lb. of steam per hour for manufacturing purposes at a pressure of 
10 lb. per sq. in. gage. Approximately how much steam will the turbine require from 
the boilers when it is operating under full load? 



195. The Importance Of Steam-turbine Lubrication cannot 
be overemphasized because steam turbines operate at such 

high speeds and are constructed 
with such small clearances that 
a slight amount of wear may 
cause disastrous results. Per- 
haps no other phase of steam- 
turbine operation is more difficult 
and has given more trouble in the 
past than has lubrication. To 
secure satisfactory lubrication, 
three fundamental requirements 
must be observed: {!) A suitable 
and high-grade oil must be used; 
see Sec. 198. (2) The oil must 
be properly supplied to the bear- 
ings; Sec. 196. (3) The purity 
and quality of the oil must be maintained; Sec. 199. 

Note. — The Functions Of An Oil In A Bearing are: (1) To form 
a film between the journal and hearing, Fig. 216, and thus to provide 
sliding between layers of the oil rather than between the metallic sur- 
faces. See the author's Steam-engine Principles And Practice for a 
discussion of the theory of lubrication. (2) To carry from the bearing 
such heat as is generated by friction in the bearing and as may flow to the 
bearing through the shaft. Sometimes, with ring-oiled bearings, water 
is circulated through the lower half of the bearing to assist in carrying 
away this heat, see Fig. 101. 

Fig. 216. — Showing how an oil film, 
L, maintains the position of a shaft in 
a bearing. The oil is assumed to di- 
vide into layers as, for example AB 
and BC. 

196. The Methods Of Supplying Oil To Turbine Bearings 
are, briefly, two: (1) Ring oiling, Figs. 75 and 217 in which a 
ring (sometimes an endless chain) is supported on the journal 
and dips at its lower part into a small reservoir of oil in the 


Sec. 196] 



pedestal. As the shaft turns, it turns the ring which thus 
carries oil to the upper part of the journal whence it is carried, 

■ -A djustin^ Scren 
'^"\\- Lock Nut 

Oil Ring- •' 'Bearing lining (L o wer) 

Fig. 217. — Bearing of the Type-6 Sturtevant steam turbine. There is an adjusting 
screw, A, in the bearing-casing cover, M. This screw when tightened down, causes 
the spherical seat, B, to grip the linings, C. The locknut, D, locks A in position. 

Electrical Connecfion To (Pongr 
From Oil floaf-^ 


Fig. 218. — Gravity oiling system used on marine turbines. {General Electric Company.) 

by the rotation of the journal, over the bearing surface. Cool- 
ing of the bearing is effected principally by radiation from the 
bearing and reservoir. Ring oiling is generally employed only 


.--Auxiliary Urn Fro m Tu rbo Pump (Startin g) 

'^ ^f ^^^^Clieck Valve- '-^yf^ , ""^ ~*t 

■ -(]H 3%" Relief Valve- ' 

Turbo Oil Pump " ^'"^ ='''-' ^ ^ 

{Starting Only)- 

'— ---^-^ A//? Suction T 
c- -/.//7e 

-J/'o/? Coc/f 

Fig. 219. — Piping diagram of the lubricating system of the Kerr turbine. The main 
oil pump is made in two parts — one to supply the governor, the other to supply the 
bearings. Hence, it is called a compound pump. The dash-dot lines indicate pipe 
lines which are below the floor. 



{///('// ////{-l 

Fig. 220. — Cross-section through Coppus turbo-blower type B, showing grease 
lubrication of ball bearings. (The grease is forced down into the cavity beside the 
ball bearings by turning the handle on the grease caps to the right, thereby forcing 
down a plunger in the cup.) 


on small turbines, up to about 300 hp. (2) Circulation oiling, 
Sec. 197, in which oil is conducted through pipes to and from 
the bearings. Circulation oihng is sometimes classified as, 
(1) gravity circulation, Fig. 218 and (2) force-feed circulation ^ 
Fig. 219, but these two classes differ only in the method of forc- 
ing the oil through the bearings — it flowing in one case by- 
gravity, in the other it is forced by a pump — the actual oil 
pressure at the bearings being small in either case. 

Note. — The Lubrication Of Ball Bearings is attained sometimes 
with grease which is supplied to them from compression cups (Fig. 220), 
and sometimes with oil which is supplied by rings as explained above 
(Fig. 50). Grease provides poor lubrication and should, generally, not 
be used except where a ball-bearing turbine is placed in a very dusty 
atmosphere — the grease then serves to keep impurities out of the bearings. 

197. The Circulation System Of Turbine -bearing Lubrica- 
tion, Fig. 221 is employed on nearly all turbines of sizes larger 
than 300 hp. and sometimes on smaller ones. The oil reser- 
voir, D, and the cooler, C, are generally provided in the bedplate 
of the turbine. In the gravity systems the reducing valve, R, 
discharges into an overhead tank. The operation of the 
system is explained below. 

Explanation. — In Normal Operation the main rotary oil pump, P, 
which is mounted on the turbine and driven from the turbine spindle, 
draws oil from the reservoir, D, through a strainer, S, and delivers it 
through the cooler, C, into a main feed pipe, M, at a pressure of 40 to 
60 lb. per sq. in. gage. This pressure is generally required to operate the 
governor. A reducing valve, R, admits oil from M to N. In N the pres- 
sure is maintained at some value from 5 to 15 lb. per sq. in. gage by valves 
R and A which are adjustable for different pressures. If the pump 
supplies more oil than is needed by the governor and the bearings, the 
excess is bypassed through the relief valve B into D. The oil in N passes 
as shown, to the several bearings where it is admitted into grooves at or 
near the tops of the bearings and is drawn between the bearing surfaces. 
Oil vents, V, prevent the accumulation of air in pockets at the bearings 
and provide a convenient means for viewing whether a bearing is receiving 
sufficient oil. From the bearings, the used and excess oil flow as shown 
by the arrows back to D. 

Should The Oil Pressure In M Fail because of clogging of the 
strainer, S, or for any other reason, the throttle valve would, in most 
turbines, be thereby automatically closed by the governor and the 


Oil Return 
To Tank- ' 

Auxi'li'ary-Pump Oil Strainer- 

Fig. 221. — Diagram illustrating the flow of oil in a circulation oiling system. For 
bearing construction see Fig. 94. {Allis-Chalmers Mfg. Co.) 



•Spindle Bushing 
Spindle Collar 
Step Bearing 

•aring Discs 

In Oil Tank 

-Turbine-driven auxiliary 
(Allis-Chalmers Mfg. Co.) 

Steam Chest 

Hoinci- Operated 
bypass Valve 

Connecfed To 

Oil Pressure 

5 y stem 

Steam To 




Fig. 223. — Pressure-operated valve 
for controlling steam supply to auxili- 
ary-oil-pump turbine on Allis-Chal- 
mers turbines. This valve automat- 
ically opens the steam supply to the 
auxiliary-oil-pump turbine and starts 
it whenever the pressure in the oil 
system falls below the normal value. 

Steam Valve 

Sec. 197] 



turbine would thus be stopped. Hence, in starting the turbine the working 
oil pressure must he attained before the turbine can be supplied ivith steam. 
For this reason, an auxiliary oil pump, T, (see also Fig. 222), driven by 
a small individual steam turbine, is supplied on each large turbine and 
is to be used in starting until the large turbine's speed is such that P can 

ffi-Governor End 

Fig. 224. — Oiling system of Ridgway turbines. Pumps A deliver oil into the over- 
head tank B. Valve C is left open until the oil level reaches D; then C is closed and the 
air above D is compressed. When the pressure in B exceeds that for which relief valve, 
F, is set (about 30 lb.), the oil flows through it and overflows at G into the lower tank, E. 
The oil which is not bypassed at F flows through the strainer, H, and thence through 
the feed-adjusting valves, /, to the bearings or through the strainer, K, to the governor. 
Sights, M, indicate the oil flow from the bearings into the return pipe, L. The used oil 
is filtered at N. Cooling water enters at O and leaves at P. A low oil pressure will 
allow F to close, which rings the alarm bell. If the oil pressure fails, the turbine should 
be stopped; the bearings will be supplied, while the rotor is stopping, by the oil in tank, 
B. The check valve, R, permits air to enter the tank in this event. The valve, F, should 
be opened only to drain the system. 

supply sufficient oil. In the smtdler turbines which are circulation- 
oiled and which do not employ oil-relay governors, oil rings are sometimes 
furnished to provide the necessary lubrication until the main pump 
attains a working speed. 

Some manufacturers equip their auxiliary-pump turbines with a 
throttle valve which is automatically controlled by the oil pressure in the 
main pipe (Fig. 223). This prevents the main turbine from coming to 


rest — which may take a half-hour or more — without a sufficient supply of 
oil in the bearings. Where the auxiliary-pump turbine is only hand- 
controlled, however, it should be started at the least indication of oil 
failure, and the main-turbine throttle valve should be closed as soon as 

Other manufacturers employ different schemes for supplying oil when 
the main pump fails. Figure 224 shows how an overhead tank may 
serve this purpose. 

198. An Oil For Turbine -bearing Lubrication Must Possess 
Certain Properties ; since the oil consumption of steam turbines 
is very low because the oil does not mix with or pass out with 
the steam or condensate but instead is largely used over and 
over again in a circulation system, it is always economical to 
use a high-grade mineral oil: (1) The viscosity should be such 
that the oil does not offer much resistance to dividing into 
layers — produce much friction — and yet the viscosity must be 
sufficiently high to insure an ample factor of safety against 
breaking down of the oil film in the bearing. An oil of high 
viscosity will cause excessive heating in the bearings and a 
consequent loss of power. Recommended viscosities are from 
130 to 310 sec. Saybolt at 100° F., although viscosities above 
200 sec. are seldom advisable; see the authors Steam-engine 
Principles and Practice for method of measuring viscosity. 
Generally, the oil of the lowest viscosity (between the limits 
given above) that will give the desired oil pressure for the 
governor, should be used. (2) Emulsification should be small; 
that is the oil should separate rapidly from water when mixed 
with it. A good comparative test is to shake like quantities 
of two oils with water in a bottle and observe the rates at 
which they separate. (3) It should he non-corrosive; a piece of 
clean polished copper inserted for 5 hr. in the oil while the oil 
is kept in a bath of boiling water should show no darkening 
or diminution of the polish. (4) It should have a flash-point 
which is not below 325° F.; oils with lower flash points are 
likely to suffer a partial evaporation in the turbine bearings 
and gradually acquire a higher viscosity. (5) It should not 
form deposits; this property can, generally, only be determined 
after a trial of the oil. 

Note. — Emulsifying And Corrosive Oils Are Particularly 
Undesirable For Turbine-bearing LuBRiCATiOtN because such oils are 

Sec. 199] 



almost certain to form a sludge or sticky compound which will clog the 
strainers, cooler tubes, and oil passages — thus impairing the lubrication 
and the cooling. 

Note. — The Following Oils Are Recommended By Various 
Turbine Manufacturers: Vacuum Oil Company's D.T.E. Light; 
Texas Company's Cetus; Atlantic Refining Company's Atlantic Turbine 
Oil, Light or Medium; Sinclair Refining Company's Cordymo; Standard 
Oil Company's Superla; Gulf Refining Company's Paramount Turbine 
Medium; Tide Water Oil Company's Turbol; Pierce Petroleum Corpora- 
tion Turbine Oils. For turbines which are subject to excessive vibration 
or which use the same oil in reduction-gear and turbine bearings (see Sec. 
203), a heavier grade should be used. 

(®X?// Pump'- 

Fig. 225. — Arrangement of apparatus in a "batch" system of oil purification. The 
dirty oil is withdrawn through valve A into the dirty oil tank below the turbine. The 
valve A is then closed and the reservoir, R, cleaned. Then valve B is opened and a 
supply of clean oil flows from the upper tank to the reservoir. Valve B is then closed 
and the turbine is ready for operation. The dirty oil is passed through the purifier and 
is pumped back to the clean-oil supply tank. {De Laval Separator Company.) 

199. The Practical Methods Of Maintaining The Purity 
And QuaUty Of The Oil Are: (1) Make-up treatment, wherein 
the oil is maintained by adding to that in the system, monthly 
or weekly, only as much oil as has been lost by leakage and 
evaporation. This, treatment is satisfactory for ring-oiled 
bearings and is sometimes employed in circulation systems. 



With this treatment, the oil should all be removed from the 
system every 3 to 6 mo. and replaced with fresh clean oil. 
If properly filtered, the oil may again be used in the bearings. 
(2) Sweetening treatment, wherein a small fraction of the oil 
in the system (3-6 gal.) is removed at regular intervals and 
replaced by fresh clean oil. During the intervals the oil which 
has been removed is thoroughly filtered and is later returned 
to the system. If sweetening is done daily, this treatment is 
very satisfactory. However, if the sweetening intervals are 
long or the amount of replenished oil too small the oil gradually 


„,-tonijnubus By-Pass (inole finite Life) 







1 n 
































1 1 ', 

_ . 'Styee t^n/nn . 















































^ Tu 






Fig. 226. — Graph showing effects of various methods of oil treatment. Graphs B 
and C might have different shapes and show better results if treatments are made with 
sufficient frequency. {Richardson Phenix Co.) 

loses its lubricating value; see Fig. 226. (3) "Batch'' 
treatment, (Fig. 225) wherein the entire oil supply is removed 
from the system at regular intervals and replaced with fresh 
clean oil. The oil which is removed is then filtered and puri- 
fied — thus making it ready for replacement into the system 
at the end of the next interval. This method of treatment 
provides very satisfactory lubrication (Fig. 226) provided the 
intervals between treatments are not permitted to become too 
great; a month say, represents good practice. A disadvantage 
of this method is that the turbine must be shut down when the 
oil is replaced. (4) "Continuous bypass'' treatment, Fig. 227, 
wherein a fraction of the oil in the system is continually passing 
through a filter, thus providing a continual '^sweetening.'' 



For turbines this method of treatment seems to be the best 
because it requires Uttle attention and gives good results. 
(5) Continuous treatment, wherein the entire quantity of oil in 
the system is filtered each time it is handled by the main oil 
pump. Although this treatment is ideal, the necessary 
equipment is costly and requires much space. Hence it is 
seldom employed. 

Slghf O^erflotY^ Turbjne Oil ReseryoiP-i^ 

Out let Td:;^\ *" Bypass For 
Sight Overflow \ p^i^j^g 

oil Pump 
oil To 


Oil Pumpy 



Fig. 227. — Illustrating one arrangement of apparatus for the 
system of oil treatment. 

continuous bypass' 

Note. — The Methods Of Purifying Oils are: (1) Precipitation and 
filtration, wherein the oil is heated, run slowly over trays, in which the 
water and heavier impurities settle out by gravity, and then is passed 
through cloth filter surfaces which remove the finer impurities. Many of 
the successful oil "filters," which are on the market, operate upon this 
principle. Their construction and operation are explained in the 
author's Steam-engine Principles And Practice. (2) Mechanical 
separation, wherein the oil is separated from the water and heavier 
entrained particles in purifiers (Fig. 228) which operate on the principle 
of the well-known cream separator wherein centrifugal force is employed 
to effect the separation. Good results are reported with these purifiers. 


They are made in different sizes to afford various capacities and require 
comparatively little power for their operation. (3) Chemical purification 
is generally necessary whenever an oil becomes acidified through use. A 
chemist should be supplied a sample of the oil end asked to recommend 
the proper treatment and, if possible, to determine the cause of the acidi- 

Sfrcf/nen . 


In let 


Helicai-Gear Dri've \: 



The Spouts May 
Readily Be Turned 
To The Mo5i 

Fig. 228. — A motor-driven centrifugal oil purifier. These purifiers are also made for 
belt or steam-turbine drive. (De Laval Separator Company.) 

fication. It should, in most cases, be possible to eliminate the trouble 
which started the acidification. 

200. The Principal Causes Of Impurity Deposits In Oils are : 

(1) Water. Where considerable quantities of water leak into 
the system, emulsion takes place, and the oil takes on a yellow- 
ish color. Furthermore it is found that a sludge or a spongy 


formation is evolved which, if permitted to remain in the 
system, will tend to clog the passages. The water generally 
leaks into the oil at the packing glands, Div. 5, or in the oil 
cooler. Water of condensation from a priming boiler wherein 
compounds are used and ''hard" cooling water are particu- 
larly troublesome. (2) Solid impurities, such as fine particles 
of rust or moulders' sand, have a marked disintegrating effect 
on oil. Where they -are present the oil assumes a dark color, 
and a ''burnt" odor. A slimy dark deposit lodges on the sur- 
faces, particularly in the cooler. Furthermore, in the presence 
of solid impurities, the oil will emulsify with very slight quanti- 
ties of water which may collect in the system and will form 
sludge. (3) Air is usually present in the oil in greater or less 
amount and will, especially if the oil temperature is permitted 
to rise above normal — say 140° F. — tend to oxidize the oil. The 
oil darkens in color, increases in acidity, and in extreme cases a 
black carbonaceous deposit develops, which may choke the 
inlet to the bearings or cause sluggish movement of the 
governor gear or may even cause it to stick. (4) Electric 
currents, in some cases, may pass down through one bearing 
pedestal, through the bedplate, and up through the other 
pedestal — a portion of the current passes through the oil, 
darkens its color, increases its acidity, and throws down a 
deposit which coats all contact surfaces and lodges particularly 
in the cooler. The deposit is of a fairly hard, brittle nature 
and of a dark chocolate color; it is very difficult to remove. 
The remedy is to completely insulate one bearing from the 
bedplate; consult the turbine manufacturer. (5) Adding new 
oil sometimes causes deposits, especially where high-viscosity 
oils are employed. 

201. Because One Function Of The Oil In Turbine Bearings 
Is To Carry Away Heat, The Oil Must Be Cooled, otherwise 
it would become too hot, lose its viscosity and become unsafe. 
Most of the heat is developed in the bearing by the friction 
between the layers of oil. Some heat also flows to the bearings 
from the steam inside the turbine casing. 

Note. — Oil Coolers, C, (Figs. 221 and 229), are generally con- 
structed of U-shaped copper tubes through which the oil (or water) is 
circulated while the outside of the tubes extends into the water space 


(or oil reservoir). It would be preferable to have straight tubes as these 
are more easily cleaned. Although most manufacturers place the cooler 
in the turbine bedplate, it is better, if possible to have it separately 
mounted so that any vibration of the turbine would not be likely to produce 
leaks at the joints. The ^pressure of the oil in the cooler should he greater 
than that of the water. This will cause leakage to occur into the water 
rather than into the oil. The oil pump should, therefore, discharge 
through the cooler rather than draw oil through it by suction. Where 
only scale-forming (hard) water is available for cooling it may give trouble 
due to deposits on the tubes. In such event it is advisable to use the 

Path ofOif? 

Circulaflng ^frif> 

Coo/er Tubes' 

Fig. 229. — The essential parts of an oil cooler. Circulating strips, as shown, are 
placed in the tubes to give the oil a whirling motion through the tubes. (Kerr Turbine 

same water over and over again by using a small cooling pond or tower to 
cool the water. 

202. The Most Desirable Oil Temperatures For Bearings 

are: (1) In circulation systems the oil comes to the bearing at a 
temperature of about 100° F. and leaves at 130 to 140° F. 
However, no trouble is usually experienced if the oil enters 
at a higher temperature and leaves at a temperature not in 
excess of 160° F. Thermometers should be fitted to indicate 
the temperature of the water entering and leaving the cooler 
and, if possible, of the oil entering and leaving each bearing. 
These thermometers should be read once every hour and the 
temperatures recorded on an engine-room log sheet. (2) In 
ring-oiled hearings the temperature of the oil in the bearings, 
if of good grade, may safely be permitted to reach 200° F. or 
even a little higher temperature. 

203. The Lubrication Of Geared Turbines (Fig. 169), since 
the service imposed on an oil in gear teeth is somewhat different 


from that imposed in bearings, is a distinct problem. As 
long as the reduction gears are perfect and run noiselessly, 
the desirable bearing oil would also be satisfactory for their 
lubrication. But, should the gears become noisy, as they 
are likely to do, a heavier oil would then be needed in the gears. 
The heavier oil would, however, not be most desirable if the 
same oil is to be used in the turbine bearings because there it 
would almost certainly be contaminated with gland water 
which would not readily separate from the oil and would give 
trouble in the gears. For these reasons, separate oil systems 
should be provided for the turbine and for the gears. 

204. The Lubrication Of A New Turbine Requires Special 
Attention because it is almost impossible to thoroughly clean 
the oiling system of all solid impurities. The impurities are 
very likely to cause deposits and hence cause trouble. The 
following procedure is therefore recommended for a new turbine. 

Explanation. — Before starting the turbine all oil tanks, pipes, the 
cooler, and the like should be thoroughly cleaned to remove such solid 
particles as dust, grit, moulder's sand, rusty scale, and cotton waste. 
Cotton waste must never be used for cleaning oiled surfaces, as it leaves 
behind small particles which tend to clog the oil pipes and the small 
spaces in the governor mechanism. A smooth, lintless cloth or a sponge 
should preferably be used. The parts should be washed first with 
kerosene and finally with clean gasoline which should be wiped dry. The 
oil should then be poured into the reservoir — not directly but through the 
sieve — and the air should be expelled from the piping with the auxiliary 
oil pump. 

After the turbine is started the strainers should be examined daily 
and, if necessary, cleaned. After a month's operation, the whole charge 
of oil should be removed from the system. The oil tanks, pipes, cooler 
and bearings should then be again thoroughly washed and cleaned. 
The system should then be filled with a complete charge of new oil. 
The oil which has been removed should be thoroughly purified and filtered 
before it is again put into the system. (It may be used as "make-up" 
oil.) This first change of oil may seem unnecessary but it will be found 
to pay in the long run; this is because a turbine requires the most care and 
attention in its early life. Later on, troubles should be rare if the oil is 
well looked after, frequently purified, and the strainers kept clean. 

205. The Care And Operation Of A Steam-turbine 
Lubrication System — see also Sec. 204 — involve: (1) Attention 
to see that each bearing is receiving oil. (2) Observation of oil 


and water temperatures, as given in Sec. 202. Abnormal tem- 
peratures will readily disclose that something has gone wrong 
in the system and will usually give an indication as to the cause 
of the trouble. In case of abnormal temperatures the unit 
should be watched very carefully and shut down as soon as 
possible. Until the machine can be stopped, more oil should 
be fed to the bearings by increasing the discharge pressure on 
the pump or by starting the auxiliary pump if necessary. (3) 
Regular treatment of the oil, according to the method (Sec. 199) 
which is employed. With regular and proper treatment a 
good turbine oil should have a life, under favorable conditions, 
of 10,000 working hours or more, or of 3,000 working hours 
under very unfavorable conditions. 

Note. — The Signs Of "Breaking Down" Of An Oil are: (1) 
Darkening in color. (2) Increased specific gravity. (3) Increased vis- 
cosity. (4) Increased acidity. (5) The 
throwing down of various kinds of de- 
posits. Although all oils are affected 
in time, unsuitable oils will break 
down much sooner than will suitable 
oils. The best oil for a system is 
therefore the one which will last the 
longest without breaking down. 

206. Some Useful Operating 
Hints On Steam-turbine Lubri- 
cation are given below: 

Do Not Pour Oil Into The Res- 
ERVoiR Except Through The 
Strainer. — It saves time in the long 
run to pour it through the strainer. 
Furthermore, since the strainer had 
to be bought, make it pay for itself. 

Take Out The Strainers (Figs. 

230 and 231) Anl Clean Them Often. 

If the strainer is on the pressure side 

of the oil pump see that the oil is 

bypassed around the strainer before 

removing it. 

Remember That Nearly All Turbines Have Some Parts Which 

Require Hand Oiling.— See that these parts get a few drops of oil 

every day. Also keep the oil cups or drop-feed oilers filled. 

E- Strainer Removed- 
Oil Bypassed 

Fig. 230. — Oil strainer which is used 
on General Electric Company's 500- 
kw. turbo-alternators. View I shows 
the normal operation. To clean the 
strainer, nut A is unscrewed. Spring 
E then forces the valve D against the 
seat F thus permitting the oil to flow 
directly to the outlet as shown in II. 
After cleaning, the strainer can be 
replaced in like manner. 

Sec. 206] 



Oil Inlet 

The Proper Oil-level For A Ring-oiled Bearing is generally- 
indicated by a scratch on the oil gage glass. See that the oil level is 

Watch Ring-oiled Bearings To See That The Rings Revolve. — 
Sometimes a ring will wear eccentric and fail to supply oil. 

Pressure Gages On The Oil System Should Be Throttled so that 
very little or no vibration is visible. 
A vibrating gage wears rapidly. 

Try To Have The Water And 
Oil Flow Through The Cooler In 
Opposite Directions ("Counter- 
flow"). — In this way less water is 
needed to cool the oil than otherwise. 

A Convenient Way To Clean 
The Oil Tubes Of A Cooler is to 
first blow them out with compressed 
air, then push through a flexible wire, 
fasten a clean cloth to one end of the 
wire, and pull the cloth through the tube. If this does not remove all 

Frame Supporting 
' e-Mi 

''•-Oil Out let Strainer dody.-'^ 

Fig. 231. — Section through oil strainer 
used on Kerr turbines. 


Va/ve 5tem 
Oil Cup 



> ^ 





7 ~ *^ 






Inlet Valve Stem 

Throttle Handwheef 
Shaft Bevel Gear, 


Valve Stem ^Sprinof 

■dfeam Inlet 

Bevel Gear 
Throttle Valve 
Stem Screw 

-Inlet Bend 

Fig. 232. — Section through throttle valve and steam chest of Allis-Chalmers turbine 
showing the oil cup, C, on the governor-operated inlet-valve stem. This cup collects 
oil which may leak through the gland from the relay cylinder and thus prevents this 
oil from "baking" on the hot valve stem. 

deposits wrap the cloth with a brass gauze and pull through again, 
will clean the tubes thoroughly. 



Provide Oil Shields, where the governor relay cylinder is located 
above the throttle valve (Fig. 232), to prevent oil which may leak out of 
the cylinder from flowing to the throttle-valve stem. If not prevented, 
the oil will ''bake" on and impair the operation of the throttle valve. 

Let The Price Of An Oil Be Your Last Consideration in making a 
selection. A "cheap" oil is expensive in the long run. 

Always Draw Off From The Bottom Of The Reservoir And, 
If Possible, After The Turbine Has Stood Idle A Few Hours. — The 
water and impurities will thus be removed. 

Take Oil Samples From The Reservoir Once A Week. — A 4-oz. 
bottle should be filled, labeled, and placed away in a safe place, k com- 
parison of these samples will often reveal troubles. 


1. Why is the lubrication of steam-turbine bearings of such vital importance? 
What three requirements are fundamental in steam-turbine lubrication? 

2. What are the two functions of an oil in steam-turbine bearings? 

3. Describe the ring-oiled method of bearing lubrication. What kind of turbines are 
most generally ring oiled? 

4. What are the two principal classes of circulation-oiUng systems? Wherein do they 

5. How are ball bearings generally lubricated? Show with a sketch how to keep 
grit out of ball bearings. 

6. Describe fully, using a diagrammatic sketch, the operation of a circulation oiling 
system. What provisions are made, in circulation systems, for supplying oil to the 
bearings in case the main oil pump should fail? 

7. State the five principal properties which an oil must possess if it is to be satis- 
factory for turbine-bearing lubrication and tell the reason for each property. 

8. What kinds of oils are particularly undesirable for turbine lubrication? Why? 

9. State the five methods of maintaining the purity and quality of an oil, describe 
each fully, and where possible draw a sketch of the apparatus required. 

10. Describe the three methods of purifying oils and give the usefulness of each. 

11. What are the five principal causes of deposit formations in oils? Explain the 
term sludge. 

12. Discuss, the construction and operation of oil coolers. How can the leakage of 
water into the oil be most easily prevented? 

13. What are the desirable and permissible working temperatures of turbine-bearing 

14. Explain fully the distinctive features of geared-turbine lubrication. What 
method of lubrication is best adapted? 

15. Why is the lubrication of a new turbine such an important matter? State what 
procedure and what precautions should be exercised. 

16. What are the three important phases of the care of a steam-turbine lubrication 
system during operation? 

17. What physical signs indicate that an oil is losing its lubricating value? 

18. State a number of lubrication "pointers" which should be observed in operating 
a steam turbine. 


207. The Three Fundamentals Of Steam-turbine Operation 

are, in the order of their importance: (1) Safetij. (2) Service. 
(3) Economy. In other words, the operator should, above 
all, endeavor to make the operation of a turbine as nearly 
free from the possibility of accident as he reasonably can; his 
next consideration should be toward eliminating the likelihood 
of a necessary shut-down; then, after these first two elements 
have been attended to, he should aim to so operate the machine 
that the economy of the plant in its use of steam is the best 
that can be attained. Safety should never be sacrificed for 
the sake of service or economy. Operating methods which 
will tend to comply with the above fundamentals are given in 
following sections. 

Note. — Some General Precautions Should Be Observed In 
Operating Steam Turbines. — The most important ones are given 
below. These precautions must be taken seriously to heart if one desires 
to obtain satisfactory operation of the turbines under his care. 

1, Understand Your Turbine Perfectly. — The preceding divisions 
were intended to familiarize the reader with the principles and usual 
construction of turbines of various types and their parts. Make sure 
that you also have the manufacturers' instructions for the turbine 
which you are to operate. Read them carefully and be sure that you 
understand them. Watch or supervise the installation of the turbine 
and be certain that you know the purpose of every piece, bolt, or nut. 
Know what is inside and out. The reasons that manufacturers have for 
doing certain things in certain ways may not always be apparent, but it is 
safe to assume that each piece has a purpose, and that the directions 
which they give have a sound basis. If a man is sent from the factory to 
acquaint you with the turbine (as is usually done with large turbines) 
ask lots of questions — he will be glad to answer them. If no man is sent, 
or if he is already gone when a question arises, write to the factory — it 
may save your life. It should always be remembered that the builders 
of the turbine know more than anyone else about the way in which that 
particular turbine should be operated. 



2. Be Sure That Everyone Concerned With The Care And 
Operation Of The Turbine Understands It Perfectly. — If the 
turbine room must be left to someone else, be sure that he is competent. 
Don't be afraid that he will "get onto " your business. Remember that 
you will get the blame if anything goes wrong. 

3. Don't Think That All Turbines Are Alike. — The fact that you 
understand one turbine does not signify that you are competent to go 
into another plant and immediately take charge of different turbines even 
if they are made by the same builder. Every turbine has its own pecu- 
liarities which must be determined by careful study. 

4. Do Not Change The Operating Conditions — steam pressure, 
superheat and vacuum — without first consulting your instruction book 
and, if the point is not covered there, writing to the manufacturer. If it 
is necessary or if you deem it advisable to change the steam pressure, 
back pressure or vacuum, extract steam, or admit steam at mixed pres- 
sures, be sure that you know what the manufacturer has to say about 
such a change. There may be small but vital details of such operation 
which you would not think of, or it may not be advisable to make such a 
change — but the manufacturer will know and will be glad to advise you. 

208. To Insure Safety In Steam-turbine Operation it is 

necessary always to observe the following points: (1) Be 
sure that the main governor operates satisfactorily. Whenever 
possible examine its parts for wear, lost motion, and sticking. 
When the unit is shut down see that the governor valve or its 
seat are not worn so that it cannot shut off the steam. (2) Be 
sure that the emergency governor operates satisfactorily. At least 
once every week or two the turbine should be speeded up to 
10 or 15 per cent, over its rated speed (according to the manu- 
facturers instructions) to insure that the emergency governor is 
in good order. Also, the turbine should always be shut down 
by tripping the emergency. (3) Keep a careful watch of the 
turbine, examining it every hour for oil temperatures, hot 
bearings and vibration (Sec. 212). (4) Be sure that the auto- 
matic vacuum breaker operates satisfactorily, if one is in the 
equipment. It is well to have a hand-operated vacuum 
breaker located near the throttle valve so that, if the auto- 
matic valve fails, the vacuum can be quickly broken by hand. 
(5) Be sure that the atmospheric-exhaust valve works properly 
and does not stick. 

209. To Insure Uninterrupted Service In Steam-turbine 
Operation the following attention is quite essential: (1) Provide 


adequate and proper lubrication, see Div. 10. (2) Always have 
a spare unit ready, if possible, to start on a moment's notice. 
When a main or spare unit is shut down for inspection or 
repairs, see that the work is done as quicklj^ as possible so that 
it will be available in case something goes wrong. The repairs 
should be so planned that any interruption of service due to 
forced shut-downs will be a minimum. (3) Make an inspection 
of the complete unit and auxiliaries at least once a year. The 
unit should be completely dismantled and every part inspected 
for wear and cleaned. In reassembling, the worn parts 
should be carefully adjusted — or even replaced where necessary . 

210. To Insure Maximum Economy In Steam-turbine 
Operation, try to: (1) Maintain the nameplate steam pressure 
and superheat, see Div. 13; this can be done by cooperating 
with the boiler-room force. (2) Maintain the nameplate 
vacuum, see Div. 13; the condenser may need frequent atten- 
tion to see that the tubes or jet nozzles are not fouled and that 
air is not leaking in. (3) Maintain the nameplate speed; 
remember that turbines are designed to operate with the best 
economy at their rated speed. (4) Operate the turbine at its 
most economical load, if possible. If more than one turbine 
must be operated to carry the total load it is sometimes best 
to have some machines run at their most economical loads 
and one to take the fluctuations whereas sometimes it is best 
to run all the machines somewhat below their most economical 
load. The most economical arrangement should be deter- 
mined by test or by reference to the individual performances 
of the several turbines and this arrangement should then be 

211. The Principal Troubles Which Arise In Steam-turbine 
Operation and which must be guarded against are: (1) Unequal 
expansion of different parts during starting; see method of 
starting in Sees. 213 to 215. (2) Water in the casing; slugs of 
water may be prevented from entering the casing by making 
the piping free of pockets and employing a separator ahead of 
the throttle valve. (3) Overspeeding ; this is guarded against 
by periodic inspection and tests of the governor and overspeed 
tripping device. (4) Excessive pressure in the casing; this is 
prevented by the atmospheric-relief valve, which should there- 


fore be kept in satisfactory condition. (5) Vibration; see 
Table 212. It is a peculiar fact that nearly all troubles 
which are experienced with steam-turbines — excepting (3) 
and (4) above — will manifest themselves sooner or later as 
vibrations. Hence, the chief duties of a turbine operator, 
while a turbine is in operation, are to carefully guard against 
overspeeding, excessive pressure, and vibrations. 

212. Table Of Vibration Causes, Remedies, and Identifying 

(Adapted from E. V. Amy, in Electric World, vol. 74, p. 1004) 


How identified 

Probable reason 

What to do 


Uniform vibration 
throughout ma- 
chine; same fre- 
quency as speed; 
becomes slightly 
less as load is ap- 
plied; intensity of 
the vibration de- 
pends on amount of 

(a) Sprung shaft. 
(6) Improperly placed balance 

(c) Displacement of balance 

weights. _ 

(d) Sediment in blades or buck- 


(e) Corroded blades. 

(/) Unequal heating of rotor 

{g) Unbalanced forces due to 

heavy distortional stresses. 
{h) Shifting of conductors on 

(z) Unequal generator air gaps. 

As soon as vibra- 
tion becomes ab- 
normal, shut 
down and investi- 
gate. Remove 
cause; rebalance 
if necessary. 


Vibration of variable 
periodicity; slight 
at no load, becom- 
ing worse as load is 

(a) Eccentric coupling. 

(6) Unequal settHng of founda- 

(c) Steam-piping strains due to 
expansion or weight. 

Check up and cor- 


Sympathetic vibra- 
tion in surrounding 
structure; vibration 
felt all over ma- 
chine and constant 
for all loads. 

(o) Improper grouting. 

{b) Non-rigid fastening of bed- 

(c) Non-homogeneous founda- 
tion resulting in unequal 

Make foundation 
solid; grout with 
lead if necessary. 


Vibration of local 
nature; greatest at 
loose bearing; rattle 
or noise when start- 
ing or slowing down. 

(a) Too much bearing clearance. 

(6) Ball joint of bearing loose. 

(c) Loose construction in built- 
up rotor. 

{d) Loose conductors on gener- 

(e) Loose coupling or bolts. 

Carefully examine 
all bearings and 
fastenings. Make 
necessary repairs. 


Abnormal vibration 
somewhat localized; 
noise varying with 
speed of machine. 

(a) Revolving buckets coming 
in contact with stationary 

ib) Insufficient casing clear- 

(c) Deflection of a diaphragm or 
disc in one stage. 

{d) Thrust-bearing troubles. 

Make repairs or 
adjustments im- 
mediately; may 
result in serious 




How identified 

Probable reason 

What to do 


Unusual noise near 
the intake; failure 
of the steam strain- 

(a) Water coming over with the 

(6) Sediment in the steam. 

(c) Faulty valve gear causing 

irregular steani admission. 

(d) Accidental closing of emer- 

gency steam valve shifting 
generator's load to other 

Make necessary re- 
pairs; test steam 
for sediment, 
acid, or salt. 


Local vibration; 
noise; heating of 
shaft or packing 

(a) Improper adjustment of 

labyrinth packing. 
(6) Packing rings too small for 


Make adjustment 
or replace old 


Heating of bearing; 
noisy operation; 
may cause damage 
to blading due to 
lowering of spindle. 

(o) Breaking down of oil film 
due to insufficient supply. 

(b) Oil supply cut off or too slow. 

(c) Poor oil (frothing, gumming, 


Improve oil sys- 
tem; keep clean 
and well filtered. 

213. In Starting A Newly Installed Turbine For The First 
Time, especially great care should be exercised. The regular 
starting operations, as described in the following sections, are 
to be followed but they should be very cautiously and slowly 
pursued. During the entire starting period there should be 
a constant watch for scant oil flow, heated bearings, blade 
rubbing, vibration trouble, and any extraordinary happening. 
If any trouble is noted, the turbine should be immediately 
stopped, the trouble corrected if possible, and then the turbine 
should be started all over again. If vibration is experienced, 
try to ascertain the cause and correct it. If a slight vibration 
begins at a certain turbine speed each time that the unit is 
started and if no cause can be found, try then to increase the 
speed and thus ascertain if the vibration will cease at a 
higher speed; the speed at which vibration begins may be a 
critical speed (Sec. 89) of the rotor shaft. Do not, however, 
try to pass any speed at which excessive vibration occurs. In 
case the cause of some trouble cannot be determined, call on 
the manufacturer for an expert erecting engineer — it is better 
to do this than to wreck your plant. It is best to permit the 
initial starting of a turbine to consume several hours and then 
to apply the load very gradually so that the machine can 
gradually ''wear itself in." It is always advisable, and will 
doubtless save money in the end, to engage an erecting 


engineer from the turbine manufacturer to erect and start 
all moderate- and large-capacity turbines. 

214. To Start A Non-condensing Turbine (Fig. 233), 
follow the manufacturer's operating instructions. It is impos- 
sible to here give any set of directions which will apply to all 

turbines. The following pro- 
cedure, will, however, apply 

-From Boilers 

^-Separator Drain 

To Atmosphere Or Low- 
Pressure System-^ 
-Throttle Valve 

in nearly all cases. 

Procedure. — (1) Start the aux- 
iliary oil pump, if the turbine has 
one. If the auxihary oil pump is 
turbine driven, first start the pump 
turbine by following instructions 2 
to 10 below. When the oil pump 
is delivering the required pressure, 
proceed with the starting of the 
main turbines as directed below. 

2. Open all valves in the drain 
pipes from the steam piping and 
turbine casing; for example, valves 
S, A, B, and C, Fig. 233. 

3. Inspect piping to see that 
the exhaust is clear and see that 
there is ample oil flowing to all 

4. Open the throttle valve, 7", 
Fig. 233, quickly but just enough 
to start the rotor spinning. 

5. Immediately, as soon as the 
rotor starts turning, trip the auto- 
matic overspeed valve by operating the hand trip lever (not shown in 
Fig. 233). This is to insure that the overspeed valve is not sticking and 
that it shuts off the steam. See that the rotor comes to rest. 

6. Reset the emergency overspeed valve. 

7. Again open the throttle valve, T, Fig. 233, to start the rotor and 
adjust the valve to give a turbine speed of about 200 r.p.m. Let the 
rotor turn at this speed long enough to insure that the turbine is thor- 
oughly warmed (3 or 4 min. on small turbines to 10 or 15 min. on large 

8. See that all bearings are receiving the proper amount of oil or that 
the oil rings are turning on the shaft. 

9. Start water flow through the cooler and bearings (if water-cooled). 
10. Gradvxilly open the throttle valve, T, Fig. 233, to increase the speed 

of the turbine. See that, at the proper speed, the governor takes con- 

V\o\z'. Pipe Drains A,&,C, And D To 
An Open 5eiver 

Fig. 233. — The principal steam 
drain piping and valves of a non-con- 
densing steam turbine. All drain pipes 
should lead from the lowest point of the 
chambers which they are to drain. 



trol. Then open T to its limit and close it one-half turn to prevent it 
from locking open. 

11. Shut down the auxiliary oil pump and see that the main pump 
keeps up the pressure. 

12. Close the valves in the drain pipes (A, B, and C, Fig. 233). If wet 
steam is used by the turbine, the drains should be left "cracked." 

13. Apply the load to the turbine gradually; see Sec. 219. 

Note. — The Rotor Should Be Spinning When It Is Being 
Warmed. — This is very important. If less steam is admitted to the 
casing than is sufficient to turn the rotor, the steam will flow through 
the casing at the top, heat the upper part of the rotor and casing, and 
thus cause unequal expansion of the rotor and casing. Later, then, when 
the rotor is permitted to turn, the distorted rotor is very likely to cause 
rubbing of the blades or a sprung spindle. But, by allowing the rotor to 
turn slowly while starting, it is warmed evenly on all sides and the cold 
air is quickly drawn from the casing. Thus unequal expansion is 

215. To start A Condensing Turbine (Fig. 234), follow the 
manufacturer's operating instructions. It is impossible here 
to give any set of rules which will apply to all turbines. Some 
manufacturers recommend starting their turbines under 
full vacuum, some under a partial vacuum (24 to 26 in.) and 
some recommend starting under non-condensing conditions. 
Whichever method is recommended by the manufacturer 
should be followed. The following procedure will be satis- 
factory in most cases. 

Procedure. — 1. Start water flow through the oil cooler and bearing 
cooling-coils and be sure that there is sufficient oil in the system. 

2. Open all drains — valves S, A,G, and D, Fig. 234 — and the turbine 
stop valve X. 

3. Start the condenser pumps; W, C, and V, Fig. 234. The dry 
vacuum or air pump may be run slowly so as to produce no vacuum in the 
condenser. Turbines with steam sealed glands may be started condens- 
ing by opening the sealing valves, K, E, and F, Fig. 234, but the vacuum 
during starting should not exceed about 25 in. 

4. Start the auxiliary oil pumps and adjust oil flow to all bearings. 

5. Open the throttle valve, T, Fig. 234, quickly to start the rotor into 

6. Immediately, as soon as the rotor starts turning, trip the automatic 
overspeed valve by operating the hand trip lever (not shown in Fig. 
234). This is to insure that the overspeed valve is not sticking and that 
it shuts off the steam. See that the rotor comes to rest. 

7. Reset the emergency overspeed valve. 



8. Again open the throttle valve, T (Fig. 234) , to start the rotor and so 
adjust the valve that the rotor turns at about 200 r.p.m. Let the rotor 
turn at this speed long enough to insure that the turbine is thoroughly- 
warmed (about 1 minute per 1,000 kw. of turbine capacity but in no 
case less than 10 minutes). 

9. See that all bearings are receiving the proper amount of oil or that 

,''Pop Valve 

Stop Valye- 

Low-Pressure ^ ^ 
.' 6/ancf 

To Hotr^ell- 

Fig. 234. 

^^Concfensafe Pump 

-Typical arrangement of piping, valves and auxiliaries for a condensing 

the oil rings are turning on the shaft. 

10. Gradually open the throttle valve, T (Fig. 234), to increase the speed 
of the turbine. See that the governor takes control at the proper speed. 
Then open T to its limit and close it one-half turn to prevent it from 
locking open. 

11. Water sealing glands may now be put into effect by turning on the 
water gradually. Then the vacuum may be raised to about 25 in. of 

12. Shut down the auxiliary oil pump and see that the main pump 
keeps up the pressure. 



13. Close the valves in the drain pipes {A, G, and D, Fig. 234). If wet 
steam is used by the turbine, the drains should be left "cracked." 

14. Apply the load to the turbine gradually; see Sec. 219. 

15. Build up the proper vacuum by regulating the condenser pumps. 

16. If a steam seal is used on the glands, close the valve — F, Fig. 234 — 
in the pipe leading to the high-pressure gland. 

Note. — To Start A Bleeder Or A Mixed-pressure Turbine, close 
the low-pressure steam valve and start as directed above for a condensing 
turbine. After the turbine is running under full load, gradually open the 
low-pressure valve. 

216. The Care Of A Turbine While It Is Running involves 
only a periodic (about hourly is generally sufficient) inspection 

DA I LY LOG SHEET FOR ?n.^. . . TURBl NE N^. . . .3 . . . . Date_ ^/-J A. | 






Water Temp. 


































































































7: 15 









































































































Fig. 235. — Portion of a turbine-room log sheet upon which entries should be made at 

inspection time. 

for unusual conditions and for taking instrument readings. 
Unusual operating conditions are generally evidenced by the 
oil temperatures, vacuum readings or by vibration or noise. 
Hence, by recording the various instrument readings hourly 
upon a log sheet (Fig. 235), troubles will generally become 
apparent as soon as they arise. If the bearing oil or governor- 
oil pressure should decrease materially, the auxiliary oil pumps 
should be started immediately. Whenever there is evidence 
of water in the turbine casing, open the drain valves to allow 
the water to escape. On turbines which have individual nozzle- 
control valves, the operator should always see to it that only the 
minimum number of nozzles required to carry the load is open. 


As a rule, it will be found that the condenser and its auxiliaries 
will require much more attention than will the turbine itself. 

Note. — Keeping Up The Vacuum On A Condensing Turbine is 
one of the most difficult tasks in connection with the operation of the 
turbine. A decreased vacuum is generally due to one of two causes: 
(1) Air leaks. (2) Fouled tubes or nozzles. To determine which of these 
factors has been the cause in any particular case is usually quite difficult. 
A scheme which is sometimes employed for finding the cause is to arrange 
some simple means for measuring the quantity of air discharged by the 
air pump. For this purpose either a gas meter may be used or a -pilot 


Pump Discharge., _^ Rece!.k?"^l°^^^^^% 

—t ■;■,.■.■■ \ uuuru 

k i^///////////////////////^ 

■ -Draff 

■^age \ f/ 

■Sharp-Edged Opening 


E-Detail Of Pi+o+ Tube 

Dry Vacuum Pump-' from Condenser 

1-Gencral Arrangement 

Fig. 236. — General arrangement and detail of pitot tube for measuring the quantity 
of air discharged by a dry-vacuum pump. The difference between the pressures in 
pipes A and 5 is a measure of the velocity of flow through the pipe C and hence indicates 
the volume of air discharged. An increase in the pressure difference, as read by the 
draft gage, indicates air leaks. 

tuhe (Fig. 236) may be inserted into the discharge pipe and connected to a 
draft gage. An increase in the quantity of air discharged will indicate 
new air leaks. 

217. Shifting Loads From One Turbine To Another, when 
they are operating in parallel, is generally effected by varying 
the governor-spring tension. If it is desired to cause a certain 
turbine to take more load, the governor spring is adjusted 
as for greater speed; (see Div. 6). This will cause it to take a 
greater fraction of the total load. Likewise, to cause a certain 
turbine to take less loads, its governor is adjusted just as it 
would be for lesser speed. In electric-power plants, the divi- 
sion of the load is generally effected by the switchboard 
operator by his operation, from the switchboard, of the motor- 
operated governor-spring adjusting device (Sec. 151). 

Note. — Working Its Field Rheostat Does Not Change The 
Power Load On An Alternator Which Is Operating In Parallel 


with another alternator; it merely changes the valve of the cross-current 
between the two machines. To adjust for minimum cross-current, 
adjust the field rheostats so that the sum of the line-current-ammeter 
readings for the two machines will be a minimum. 

218. To Stop A Turbine which is operating under load, it is 
customary to gradually decrease the load on the turbine, 
before shutting off the steam supply. This procedure is not 
essential, however, as no harm will result to the turbine if the 
steam supply is shut off while the machine is under load — 
harm may, however, result under some conditions to the 
machine which the turbine drives. The following procedure 
in stopping a turbine will apply in nearly all cases. 

Procedure. — 1. Start the auxiliary oil pump. 

2. Gradually decrease the load on the turbine by varying the governor- 
spring tension; (Sec. 151). When the load is reduced to about one-tenth 
of full load, reduce the vacuum to 24-26 in. by opening the vacuum 
breaker valve. Remove the entire load if possible. 

3. Pull the trip lever to close the emergency-governor valve and allow 
the turbine rotor to come to rest. See that the bearings are receiving oil 
while the rotor is stopping. 

4. Stop the auxiliary oil pump. 

5. After about 15 minutes stop the condenser pumps. This will 
insure that all water vapor is drawn from the turbine casing. 

6. Open all drains and leave them open until the turbine is started 

7. Close the turbine stop valve, X(Fig. 234), and open a drain between 
it and the throttle valve, T. This will prevent steam from blowing past 
the throttle valve and tending to cause leakage. 

219. To Apply The Electrical Or Mechanical Load When 
Starting A Turbine, the following instructions will be found 
of value. It is assumed that the turbine which drives the load 
has been started as outlined in Sees. 214 and 215. 

1. To Start A Single Alternator. — (a) Start the exciter and adjust 
for normal voltage. (6) Turn the generator field rheostat so that all of 
its resistance is in the field circuit. Close the field switch, (c) Adjust 
the rheostat of the exciter for normal voltage. Slowly increase the 
voltage to normal by cutting out the resistance of the field rheostat. 
{d) Close the main switch. 

2. To Start An Alternator To Run In Parallel With Others. — 
(a) Adjust the exciter voltage and close the field switch, the resistance 


being all in, as described above, (6) Adjust the generator field resistance 
so that the generator voltage is the same as the bus-bar voltage, (c) 
Synchronize the generator with the bus-bars — see the author's American 
Electrician's Handbook. Close the main switch, {d) Adjust the 
field rheostat until cross currents are a minimum and adjust the governor 
of the turbine so that the load will be distributed, as desired, among the 
operating generators. 

3. To Start A Direct-current Generator. — (a) Before starting 
the turbine close the field switch and see that the entire rheostat resistance 
is in the field circuit. Then bring the machine up to speed. (6) Cut out 
field resistance to raise the voltage to the rated value or, if parallel 
operation is desired, to the voltage of the bus-bars, (c) Close the main 
line switch, {d) Adjust the load on the generator by varying its field 

4. To Start A Centrifugal Pump. — (a) Before starting the turbine 
prime the pump and close its discharge valve. This will permit starting 
under fractional load. Then bring the machine up to speed. (6) Open 
the discharge valve gradually to put load on the pump. See also the 
author's Steam Power Plant Auxiliaries And Accessories. 

220. To Take The Load Off Of A Turbine in stopping it, the 
procedure is generally the reverse of that which is performed 
in starting up and applying the load. To avoid misunder- 
standing the following instructions are given : 

1. To Stop A Single Alternator. — (a) Decrease the field current 
by turning in all of the field-rheostat resistance. (6) Stop the turbine, 
(Sec. 218). (c) Open all switches and stop the exciter. 

2. To Cut Out An Alternator Which Is Running In Parallel 
With Others. — (a) Partly close the turbine throttle valve so that the 
load on the generator is reduced. (6) Open the main switch. Do not 
open the field switch before opening the main switch, (c) Stop the 
turbine, {d) Open the field switch and stop the exciter. 

3. To Stop A Single Direct-current Generator. — (a) See that all 
motors are disconnected from the fines. (6) Stop the turbine, (Sec. 218). 
(c) Turn all rheostat resistance into the field circuit, {d) Open the main 

4. To Stop A Direct-current Generator Operating In Parallel 
With Others. — (a) Reduce the load as much as possible by throwing all 
resistance into the field circuit with the field rheostat. (6) Throw off the 
load by opening the circuit-breaker, if one is used; otherwise open the 
main switch, (c) Stop the turbine, (Sec. 218). 

5. To Stop A Centrifugal Pump. — (a) If the pump is operating in 
parallel with others, close the discharge valve. (6) Stop the turbine, 
(Sec. 218). 


221. Regular Inspections Of Steam Turbines Should Be 
Made. — The object of such inspections is to find the source of 
some possible trouble before the trouble actually shows itself. 
Since all turbines operate at high speeds and with only rotating 
motions, slight amounts of wear will not give warning as by 
knocks or the like, but will increase until some serious damage 
occurs — such as the rubbing of blades or the burning out of a 
bearing. To forestall such damage, the following inspections 
are recommended. 

1. Hourly Inspections. — Hourly readings should be taken of the 
temperatures and pressures of the oil at various points in the system, 
the temperatures of the circulating water and condensate, the vacuum in 
the condenser, the pressure and superheat of the supply steam, steam 
pressures in various stages of the turbine, load on the turbine, and other 
like quantities. These readings, together with any unusual noise or cir- 
cumstances, should be recorded on a log sheet (Fig. 235) which is kept for 
the purpose. Irregularities in any of these readings will immediately dis- 
close some approaching trouble. 

2. Monthly Inspections. — At least as often as once a month, a test 
should be made on the emergency governor by gradually increasing 
the speed of the turbine above normal rated speed to that at which the 
governor should shut off the steam supply. If the governor operates, the 
speed should be recorded. If the governor does not operate it should be 
adjusted or repaired. The steam strainer should be inspected, cleaned if 
necessary, or if in poor condition it should be replaced. The alignment 
of the unit should be checked very carefully. In some installations, 
measurements are made each month for possible settling of the founda- 
tion. The adjustment of the thrust bearing should also be checked, 
(Div. 5). 

3. Yearly Inspections. — Once each year the entire unit should 
be dismantled, cleaned, and all parts inspected for wear. The steam 
passages should be carefully examined for erosion. Badly worn valves, 
nozzles, or blades should be replaced if possible. It is to be expected that, 
after a number of years of service, the parts which are subjected to the 
action of steam flow will be worn quite badly. In such cases, new parts 
should be obtained from the manufacturer. When the parts are again 
assembled, all bearings should be adjusted (see Div. 5) so as to obtain the 
proper clearances and ahgnment. 

222. The Maintenance Of Steam Turbines, aside from the 
periodic inspections. Sec. 221, involves only: (1) Keeping up 
the purity and quality of the oil; this is treated fully in Div. 10. 
(2j Making adjustments and replacements; the bearings should 


always be so adjusted that the ahgnment and clearances are 
correct; worn bearings, which will scarcely ever be found if the 
lubrication and alignment are carefully attended to, may be 
rebabbitted (Sec. 97) or replaced; badly worn nozzles, blades, 
or valves should be replaced. 

Note. — The Repair Of Broken Blading should not be atten pted by 
the turbine operator. Such repairs should be made by the manufacturer 
of the turbine, because it is essential that the repaired blading bg tested 
for strength and balanced before being put to service. Sometimes, if a 
unit on which some blades have broken cannot be spared from service 
for some time, a temporary repair can be effected by cutting out all of the 
blades which remain in the rows from which some have been los.*:. This 
will restore the balance of the rotor and will permit running the turbine at 
a slightly reduced capacity and with but a slight loss of efficiency. 
Later, when the unit can be spared and the manufacturer is ready to make 
the repairs, the rotor may be shipped to his factory. 

223. If A Turbine Will Not Carry The Load Which It 
Should, the cause is most probable one of the following 
{Terry Instruction Book) : 

1. Excessive Load. — (a) Overloaded driven machine. (6) More 
power required than the turbine was built to develop, (c) Wear of driven 
machine has lowered efficiency, requiring more power. 

2. Plant Conditions. — (a) Steam pressure at the throttle less than 
that stamped on the nameplate. (6) Turbine designed for superheat 
but run on saturated steam, (c) Turbine designed for dry steam, but 
very wet steam used, {d) Back pressure in casing greater than specified, 
(e) On condensing turbine, vacuum is low. 

3. Turbine Adjustments. — (a) Hand valves closed that should be 
open. (6) Governor closes valve before normal speed is reached, (c) 
Valve improperly set, (see Div. 6). (rf) One or more jets plugged, (e) 
Clogged strainer in steam line. (/) "Lap" or "Clearance" wrong, (see 
Div. 5). ig) Buckets worn by wet steam or otherwise. If so, describe 
conditions to manufacturers and they will advise, ih) Parts binding or 
rubbing. {%) If turbine has been taken apart the wheel may be on back- 
wards or, in a multi-stage turbine diaphragms or wheels interchanged. 

224. If The Steam Consumption Of A Turbine Becomes 
High, the probable causes {Terry Instruction Book) are: 

1. The Same Causes As For Insufficient Power, (Sec. 223). 

2. Hand Valve Control. — (a) Keep as many hand valves closed as 
load conditions will allow, and thus keep the pressure in the steam ring 
as high as possible, to get the best use of the steam pressure available. 


(6) Do not run with hand valves "cracked." Keep them either open or 
shut, (c) Inspect hand valve seats. Leakage here will cause loss when 
valves are closed. 

3. If The Turbine Runs Below Speed, the water rate will be 
increased and the capacity decreased. In the case of pumps running 
from a pressure governor, however, the overall efficiency of the unit is 
benefited by running at reduced speed when lightly loaded, on account 
of reduced pump losses. 

225. When Writing To The Factory For Advice, the follow- 
ing information should be given {Terry Instruction Book): 

1. When writing to the Terry Steam Turbine Company, in regard 
to the power or economy of a turbine, please read the above tabulation 
(Sees. 223 and 224) and so far as possible advise us on the various points 

2. Take readings as follows with the turbine running under load, 
repeating for several loads if possible: (a) Steam pressure in the steam 
line between the throttle valve and the turbine. (6) Superheat or 
moisture in the steam, (c) Steam pressure in the steam ring, (d) Num- 
ber and position of hand valves open and closed, (e) R.p.m. of turbine. 
(/) Back pressure at the turbine exhaust, {g) Load on driven 
machine, if measurable. 

3. Give all information on the name plate of the machine especially the 
serial number. 


1. State the three fundamentals of steam-turbine operation in the order of their 

2. What precautions should be observed if successful operation of a turbine is to be 

3. List about 5 points which affect the safety of a turbine's operation. 

4. What factors tend for uninterrupted service in turbine operation? 
6. How should a turbine be operated to insure maximum economy? 

6. State the five principal troubles which are likely to arise in the operation of a 
turbine. How are they guarded against? 

7. What are the eight principal causes of turbine vibrations? How would you dis- 
tinguish which is the cause in any particular case? 

8. What special precautions should be exercised in starting a newly installed turbine 
for the first time. 

9. Give the steps required in starting a non-condensing turbine. Illustrate with a 

10. Should the turbine rotor be turning when the steam is turned on to warm it? 

11. State the procedure of starting a condensing turbine. Illustrate with a sketch. 
Is the turbine started under a vacuum or non-condensing? 

12. What special procedure should be followed in starting a bleeder or a mixed- 
pressure turbine? 

13. What care does a turbine require while it is running? 

14. What are the two principal causes of a gradual decrease in the vacuum on a 
turbinfi? How may an operator distinguish the actual cause in a given case? 


15. Make a sketch of and describe the pitot-tube method of measuring air discharge. 

16. Should a turbine be stopped with the load on or after removing the load? Why? 

17. State the usual procedure of stopping a turbine. 

18. Describe the methods of applying the load to a turbine with (o) a single alter- 
nator, (6) an alternator which is to run in parallel with others, (c) a direct-current 
generator, {d) a centrifugal pump. 

19. Explain how to take off the load from a turbine which is driving (a) a single 
alternator, (fe) an alternator which is running in parallel with others, (c) a single direct- 
current generator, (d) a direct-current generator in parallel operation with others, (e) a 
centrifugal pump. 

20. What are the purposes of making regular inspections of a turbine? 

21. Explain what should be done at each hourly inspection. 

22. Explain what should be done at each monthly inspection. 

23. Explain what should be done at each yearly inspection. 

24. What are the essential points in the maintenance of a steam turbine? 


226. The Purpose Of Testing A Steam Turbine For Per- 
formance is to obtain data whereby the performance values, 
or heat economy, may be computed (Sec. 240). The perform- 
ance values which are computed from the results of the test 
may be used in determining: (1) How nearly the 'performance 
of the turbine approaches or exceed0hat which was guaranteed 
hy the manufacturer. A test for this purpose is called an accep- 
tance test. (2) Whether or not an old turbine is operating at 
its maximum efficiencii.^ (3) The comparative performance of 
two or more prime movers. (4) The overall economy of the power 
plant. Various methods of testing steam turbines are described 
hereinafter in this division. 

Note. — The Conditions Under Which A Test Is Made Should Be 
Governed By The Object Of The Test. Turbines are usually sold 
under a guarantee (Sec. 285) which is based upon certain operating condi- 
tions, such as the initial and final conditions of the steam, speed of rota- 
tion, and load. Consequently, if the results of a test are to be used in 
comparing the actual operating performance with the guaranteed per- 
formance, the conditions under which the test is made should conform 
as nearly as possible to those specified in the guarantee. However, if 
the object of the test is to compare the performances of two prime 
movers on an economic basis, the test of each should be made under the 
conditions for which it was designed. Then, a correction (Sec. 268) 
should be made to reduce both performances to the same, comparable, 
basis. In testing a turbine to determine the overall economy of a power 
plant, the conditions under which the test is made should, as nearly as is 
possible, conform to the conditions under which the plant normally 

227. The More Important Data Obtained In Testing A 
Steam Turbine are: (1) Condition of the steam entering the 
turbine. (2) Condition of the steam at the turbine exhaust. 
(3) Power output of the turbine. (4) The quantity of steam 



consumed hy the turbine. (5) The speed of rotation of the turbine 
shaft. Various methods of obtaining these data are described 
in following sections of this division. 

228. The Duration Of A Steam-turbine Test should ordi- 
narily be from 3 to 5 hr. However, the object of the test 
(Sec. 226) may render it desirable to extend the test period 
over a longer time. A test over a period of less than about 
3 hr. cannot be relied upon for accurate results. The readings 
of the various instruments should be made and recorded 
''Fig. 245) at intervals of not more than 30 min. 

Note. — The Duration Of A Steam-turbine Test As Specified By 
The A.S.M.E. Code is quoted below. Where practicable, this speci- 
fication should be followed: "A test for steam or heat consumption, with 
substantially constant load, should be continued for such time as may be 
necessary to obtain a number of successive hourly records, during which 
the results are reasonably uniform. For a test involving the measure- 
ment of feed water for this purpose, 5-hr. duration is sufficient. Where a 
surf ace_condenser Js used, and the measurement is that of the water 
discharged by the condensate pump, the duration may be somewhat 
shorter. In this case, successive half-hourly records may be compared 
and the time correspondingly reduced. | When the load varies widely at 
different times of the day, the duration should be such as to cover 
the entire period of variation." 

229. The Apparatus And Instruments Which Are Required 
For Testing A Steam Turbine depend upon the object of the 
test (Sec. 226), and upon the local conditions and arrangement 
of the plant. In general, however, those instruments which 
are Hsted in the A.S.M.E. Code (Sec. 248) should be available. 
All instruments which are used should be accurately calibrated 
according to the rules of the A.S.M.E. Code before and after 
each test. Then, the observations should be corrected for 
any errors which may be noted in the instrument readings. 

230. The Condition Of The Steam Entering The Turbine 
Is Determined by: (1) The pressure, in pounds per square 
inch, which is read from a pressure gage, P, Fig. 237. (2) 
The temperature of superheat or the quality. The temperature 
of the gteam is determined by a thermometer (Fig. 238 and 
T, Fig. 237). Then, from a steam table, determine the 
temperature of saturated steam at the pressure indicated by 

Sec. 230] 



.■■Live Steam From Boiler 


Two Waffmefers 
For Measuring PoYV&r 
■In A yPhase System 

Fig. 237. — Illustrating arrangement of apparatus for testing a small-capacity steam 
turbine driving a three-phase generator and exhausting into a surface condenser. 


Fig. 238. — Showing method of obtaining the temperature of steam which is flowing 
in a pipe. (The length of the well should be such that the bulb of the thermometer 
will be at about the center of the pipe.) Unless the thermometer which is used is one 
that is graduated for the specific "immersion," its readings should be corrected for 
"stem exposure; ' see ine author s I'ractical Heat. 


■Sfeam Fbw Mefer- -> 




Fig. 239. — Illustrating location of apparatus for testing a steam turbine which has a 
back-pressure-turbine-driven jet condenser, J . The steam consumption is determined 
by a steam flow meter, 5, or by water meters, M. 


the pressure gage, P. If the temperature as read from the 
thermometer {T, Fig. 237) Js^noreJ^han^^bout^lO^R^^^^ 
that found in the steam table corresponding to the reading of 
J^jt is reasonably certain that the steam is superheated.l The 
temperature of the superheat will then be the difference 
between the temperature as read on the thermomet er and ^tbgi 
temperature of the steam as obtained from tjie stea m table. 1 
If_the^ difference between the thermometer reading and the 
temperature of saturated steam as obtained from the steam 
table is less than 10 °F., the steam may be wet, and its quality 
sEould be determined jDy a calorimeter, C, Fig^_237. 

Note. — The Location Of The Instruments For Determining The 
Condition Of The Steam Entering The Turbine should be as near to 
the steam-inlet flange as is practicable (see Figs. 237 and 239). The 
throttle, V, Fig. 237, should be wide open during the test. 

231. The Property Of The Steam At The Turbine Exhaust 
Which Must Be Determined Is The Temperature. However, 
as stated below, both the temperature and the pressure of the 
exhaust steam are usually noted. The temperature of the 
steam is determined by inserting a thermometer (Fig. 238) in 
the exhaust pipe of the turbine. This thermometer (E, Figs. 
237 and 239) should be located as near as is practicable to 
the turbine exhaust flange. 

Note. — If The Exhaust Pressure Or The Condenser Pressure 
Is Determined By A Pressure Gage — of either the Boudon-tube or 
mercury-column type — the reading of this gage should be recorded 
as referred to a barometric pressure of 30 in. of mercury. That is, 
if during the test the barometric pressure is 29.5 in. of mercury, and the 
pressure gage indicates a condenser pressure of 27.5 in. of mercury, the 
condenser pressure referred to a SO-in. barometer = 30 — (29.5 — 27.5) 
= 30 — 2 = 28 in. of mercury. Thus, the condenser pressure as referred 
to a 30-in. barometer results in the pressure which would be indi- 
cated by the vacuum gage if the atmospheric pressure were 30 in. of 
mercury. For accurate results, a mercury-column pressure gage should 
be used for determining the exhaust and condenser pressures. The baro- 
metric pressure should be determined by a barometer which is located 
near the pressure gage. If no barometer is available for reference, the 
barometric reading may be obtained from the local Weather Bureau. 

232. The Power Output Of The Turbine May Be Deter- 
mined : (1) Mechanically, by a brake, such as a prony brake, or 


a water brake. For methods of obtaining the power output 
by a brake see the author's Steam-engine Principles and 
Practice. (2) Electrically, by measuring the electrical 
energy or the power output of the driven generator. This 
method, which is described in the following sections, is practi- 
cally always employed in testing turbo-generators. 

233. The Power Output Of A Turbo -generator May Be 
Determined Electrically At The Generator Terminals by 
wattmeters, ammeters and voltmeters, or watt-hour meters. 
Whichever instruments are used should be of the portable type, 
and should be so screened that they will not be affected by any 
stray magnetic fields. If the load remains practically constant 
throughout the test, the use of wattmeters (TF, Fig. 237) will 
generally result in greater accuracy than will the use of a watt- 
hour meter. However, if during the test, the load fluctuates 
materially, a watt-hour meter should be used. Then, the 
average power output, in kilowatts, may be determined by 
dividing the number of kilowatt-hours, as indicated by the 
watt-hour meter, by the number of hours duration of the 
That is: 

(31) Av. kw. power output = 

Kw.-hr. generated during test ., ., v 

Hours duration of test 

Example. — If during a certain test, of 4-hr. duration, 4,876 kw.-hr. of 
energy are generated, what is the average power developed during the 
test? Solution. — Substitute in For. (31): Av. kw. power output = 
(Kw.-hr. generated during test) /{Hours duration of test) = 4,876 -^-4 = 
1,219 kw. 

Note. — In Measuring The Power Output Of A Turbo-alter- 
nator Electrically, it is preferable that the load on the alternator be 
as near unity power factor as is possible. The reason for this is, that if 
the power factor of the load is unity, the error which would otherwise be 
caused by phase displacement in the instrument transformers will 
be obviated. A load at practically unity power factor may be obtained 
by connecting the generator to a water rheostat. If a three-phase 
alternator is operating under an inductive load — power factor less than 
unity — the proper balancing of the load on each of the three phases 
should be checked by the station ammeters and voltmeters. For detailed 
instructions for measuring the electrical output of generators, see the 
author's Steam-engine Principles and Practice. 


234. The Power Output Of A Generator Should Be Deter- 
mined As The *'Net Watts" Output.— That is, the power 
required for excitation should be recognized in determining 
the power-output value of the generator. Thus, if the 
generator is self-excited (direct-current generator) or if 
the exciter is direct-connected to the turbo-generator shaft 
(as it is on some turbo-alternators) the energy for excitation 
need not necessarily be considered. However, if the generator 
is separately excited, the power, in watts, which is supplied to 
the generator for excitation must be measured. Then, to 
obtain the net power output of the generator subtract the 
power input required for excitation from the power output as 
measured at the generator terminals, see colums 5, 6 and 7 
(Fig. 245). That is, for a separately-excited generator: 

(32) Net kw. output = (Kiv. output at terminals) — 

(Kw. excitation) (kilowatts) 

235. The Quantity Of Steam Consumed By The Turbine Is 
Generally Determined By One Of The Following Methods : 

(1) By measuring the condensate. (2) By measuring the feed 
water. (3) By a steam-flow meter. The first method — that of 
weighing the condensate — will, generally, result in greater 
accuracy than will anj^ of the other methods. Consequently, 
where practicable, it should be used. Each method is 
described in a following section. 

236. To Determine The Quantity Of Steam Consumed By 
A Turbine By Measuring The Condensate Discharged From 
A Surface Condenser, the water which is discharged by the 
condensate pump is generally piped (Fig. 237) to tanks, R, 
which set on weighing scale platforms and is there weighed. 
Thus, by weighing the condensed steam which is discharged 
durin^^a certain number of hours, and dividing the total 
weight by the number of hours,Jthe total steam consumption^ 
in pounds per hour, results. That is: 

(33) Total lb. per hr. steam consumption = 

Lh. condensate discharged during test 


Hours duration of test 
This method of determining the steam consumption is only 



practicable where the turbine is operated (Fig. 237) in con- 
junction with a surface condenser. The arrangement of tanks, 

'Scale Platforms- 

--; ^.r^rVerfical Sump 



H - E. r e V a t » o n 

Fig. 240. — A convenient arrangement of tanks and piping for weighing the condensate 
from a surface condenser. 

piping, and scales for weighing the water, and corrections 
which must be made are treated in the notes below. 

Note. — A Tank And Piping Layout For Weighing The Con- 
densed Steam Which Is Discharged From The Surface Condenser 


is shown in Fig. 240. The discharge pipe from the condensate pump is 
tapped at T (Fig. 240-7) and an arrangement is made as shown for by- 
passing the water through the tanks A and B. After one tank, A, has 
filled, the water from the condensate pump may, by means of the quick- 
opening three-way valve, D, be diverted into the other tank, B. Then 
while B is filling, the water in A may be weighed. After it has been 
weighed the quick-opening valve in the large outlet pipe, C, is opened, so 
that by the time B is full, all of the water that was in A has been weighed 
and discharged into the reservoir tank, E. The water is removed from 
E by the vertical motor-driven centrifugal pump, P. The dimensions of 
the tanks as shown in Fig. 240-7/ should provide sufficient capacity for 
testing a 5,000-kva. turbine. In the event that scales are not available 
for weighing the water which is discharged from the condenser, its weight 
may be computed by the following formula: 

(34) W = AhD (pounds) 

Wherein: W = the weight, in pounds, of the water in the tank. A = 
the cross-sectional area of the tank, in square feet, h = depth, in feet, 
of the water in the tank. D = the density of the water, in pounds per 
cubic foot, at the temperature of the water in the tank. To obtain D, 
it is necessary to measure the temperature of the water in the tank. 
Then from a table of densities of water (this is given in most steam 
tables), find the density in pounds per cubic foot at the measured 

Note. — In Measuring The Condensate From A Surface Con- 
denser, The Amount Of Leakage Of Either From The Condenser- 
circulating-water Passages Or From Other Sources (Sec. 248) 
Must Be Determined And Proper Corrections Made. — One method 
of determining the condenser leakage is to raise the vacuum in the 
condenser to the operating value and, with the throttle (F, Fig. 237) 
closed, determine the amount of water which is discharged by the con- 
densate pump. This test of condenser leakage should be continued for a 
period of at least 2 hr. If the leakage test results in an appreciable 
amount of water being discharged from the condensate pump, the leaks 
in the condenser should be located and repaired before proceeding with 
the turbine test. This is because that, when the turbine is operated at 
full load, the leakage will be much greater than it was when the leakage 
test was made with the throttle closed. There are other methods of 
determining the condenser leakage at full load, such as by chemical 
titration or by electrical resistance, but they will not be described herein. 
Any water leakage into or out of the condenser from the turbine or pump 
glands must be determined and proper correction made therefor. 

237. To Determine The Quantity Of Steam Consumed By 
A Turbine By Measuring The Boiler Feed Water (F g. 241), 
the water may be piped from the feed-water heater, H, to 


tanks A and B, which are supported on weighing-scale plat- 
forms, where it is weighed. The water is, after weighing, 
emptied into the reservoir, R. From R it is pumped to the 
boiler by the boiler feed pump, P. The water level in the 
boiler, as indicated by the water gage thereon, should be 
the same at the end of the test as it was at the beginning. 
Then, by deducting the leakage (see note below) from the 
total weight of the water pumped into the boiler during the 
test, the steam consumption for the duration of the test results. 

Note. — In Determining The Steam Consumption By Measuring 
The Boiler Feed Water, The Leakages For Which Corrections 


•From Boiler 

, Turbine- 

■Generator po 

^- -Live Steam 


Open Feed' loafer 
Heater- , 

, Back- Pressure Turbine- 
■Drlyen Boiler-Feed Pump 

■Plat Form Scales-^ H-->;nng 

V^/^j/// ^^^^ '''^ V'^^''' '^'/ ^^^^/^/7//// 

Circulating J 
Pump --' 


^^^^ v^^^m^^Mm. ' ■ ~ -Reservoir 
Fig. 241. — Showing arrangement of tanks for weighing boiler feed water. 

Must Usually Be Made Are : ( 1 ) The leakage of water which occurs in the 
boiler feed pump, and in the pipes between the reservoir {R, Fig. 241) and 
the boiler. The amount of this leakage may be determined by closing 
off all the feed valves at the boiler, "running the pump, P, for about 15 
min., and noting the quantity of water which has disappeared from the 
supply tank, R. In making this test, a pressure gage should be placed in 
the pump discharge to guard against a dangerous water pressure in the 
pipe. During this leakage test, the reading of this pressure gage should 
be approximately that of normal operation when the feed valves are open. 
(2) The leakage of steam from the boiler, and from the connections and valves 
between the boiler and the turbine. This leakage may be determined by 
shutting off the feed-water supply, and by breaking and blanking off all 
branch connections to the steam line which connects the boiler to the 
turbine. Then, by means of a slow fire, maintain the steam pressure at 
the same pressure which is to obtain during the test. This pressure 


should be maintained for a period of at least 2 hr., and the water level in 
the gage glass should be noted at about 15-min. intervals. The amount 
of steam which has leaked out may be computed by the amount of the 
decrease in the water level as shown by the water gage. For more 
detailed instructions concerning these leakage tests, see the A. S. M. E. 
Test Code (Sec. 248). 

238. A Steam-flow Meter May Be Used To Determine The 
Steam Consumption Of A Turbine. — The meter should be 
connected into the high-pressure steam line (S, Fig. 239) near 
the turbine. It should be calibrated in that place with 
approximately the same temperature, pressure and steam flow 
as will obtain during the subsequent turbine test. A steam- 
flow meter cannot generally be depended upon for an accuracy 
of more than about 2 per cent. Where accurate test results 
are desired it should not be used. 

Note. — Water Meters May Be Used To Measure The Water 
For Determining Steam Consumption, either in the boiler feed-water 
line or in the condensate-pump discharge. The condensate in a jet 
condenser, J, may be determined (Fig. 239) by metering the injection 
water ^nd the discharge water, and then taking their difference. Water 
meters, M, when used, should be frequently calibrated in place. They 
cannot be depended upon for accurate results. 

239. The Speed Of Rotation Of A Turbine Rotor Is Gen- 
erally Determined By A Tachometer (Figs. 242 and 243). If 
the power output is to be deter- 
mined electrically (Sec. 233), the '''''- ^''"'■ 
only purpose of the tachometer 
is to insure that the rated speed 
is maintained constant through- 
out the test. For a turbo-alter- ^ ^ ^ ^^^^'' ^.'"'- ^ 

i*iG. 242. — Vibrating-reed tachom- 

nator which is equipped with a eter. {Jas. g. Biddu Co.) 

frequency meter, the tachometer 

may be dispensed with. However, if the power output is to 
be determined by the brake method (Sec. 232) an accurately- 
calibrated tachometer is essential. 

240. The Various Terms And Efficiencies Which Are 
Generally Used To Express Steam-turbine Performance 
Values (Sec. 226) are discussed and explained in following 


Sees. 241 to 245. The terms are: (1) The water rate, which is 
expressed as the number of pounds of steam required to 
generate a kilowatt-hour or a horsepower-hour of energy. 
A water rate graph is shown in Fig. 244. If the turbine is 
used to drive a generator, the water rate is usually expressed in 
pounds of steam per kilowatt-hour. If used to drive a pump, 
compressor, or the like, the water rate is usually expressed in 
pounds of steam per brake horsepower-hour. Whenever the 
water rate of a turbine is given as its performance value, the 


baseplate--'' Pulley-' 

Fig. 243. — Electric tachometer. (The tachometer consists chiefly of a direct-current 
magneto, M, and a voltmeter, V. The pulley, P, is driven by belt from the shaft the 
speed of which is to be measured. Since the magnetic field of M is produced by per- 
manent magnets, the voltage which it generates will be proportional to its speed. 
Hence, the scale of F can be calibrated to indicate revolutions per minute directly.) 

steam conditions at inlet and exhaust should also be given; 
unless the steam conditions, are stated, the water rate is a 
very indefinite performance value. (2) The number of heat 
units required to develop one unit of mechanical or electrical 
energy, which is expressed as the number of British thermal 
units per kilowatt-hour or per brake horsepower-hour. (3) 
The net mechanical work developed by one heat unit, which is 
expressed in foot-pounds of net work per British thermal unit. 
(4) The thermal efficiency, expressed as a percentage. (5) 
The Rankine cycle ratio expressed as a pe rcen tag^,^ The 
example given below is merely to illustrate the method of 
computing the above performance values from assumed test 
data of a turbo-alternator, and is not intended to represent 
the performance of any particular machine. 

Sec. 240] 



Example. — The half-hourly observations of a full-load test on a 
10,000-kva. turbo-alternator are as recorded in Fig. 245. Compute the 
following performance values: (a) The water rate. (6) The number of 
British thermal units consumed -per kilowatt-hour, (c) The number of foot- 





3 12,000 




Q. 6,000 







_ 2,000 






Zl i^ 



400 600 

In Kilowatts 



Fig. 244. — Graph showing total steam consumption and the water rate of a 1,000-kw. 
steam turbine. (Dotted lines show the guaranteed consumption. Full hues show con- 
sumption, as determined by official test. The graphs are for a 1,000-kw., 3,600-r.p.m., 
turbine, for the City of Grand Rapids, Michigan, operating under the following steam 
conditions: Dry saturated steam at a pressure of 140 lb. per sq. in., gage, and a vacuum 
of 28 in. of mercury, referred to a 30-in. barometric pressure. Allis-Chalmers Mfg. Co.) 

pounds of net work developed per British thermal unit, (d) The thermal 
efficiency, (e) The Rankine cycle ratio. 

Solution. — The averages of the half -hourly data readings, as recorded, 
are computed and entered in the last line of Fig. 245. The values of 
Hi and H2 (at bottom of Fig. 245) are 1,252 and 894 B.t.u. respectively, 
determined from a steam chart (Fig. 15) on the basis of the supply steam 


at a pressure of 150.8 lb. per sq. in. gage and a superheat (Sec. 230) of 
100.4° F. at the throttle, and the exhaust steam at a temperature of 
92.8° F. The value of ha, 61 B.t.u., as determined from a steam table, 
is the number of British thermal units in 1 lb. of water at the temperature 
(92.8° F.) of the exhaust. 

(a) The Water Rate May Be Determined by the following formula : 



y^K = 


(lb. per kw.-hr.) 

Wherein: Wk = the weight of steam, in pounds, required to develop 
1 kw.-hr. W = the total weight of steam, in pounds, consumed by the 
turbine for the duration of the test, t = the duration of test, in hours. 
Px = average net power, in kilowatts, developed by the turbine. From 
Column 9, Fig. 245, W = 436,800 Ih. From Column 1, Fig. 245, t = 3 







In. Gage 


•K Or . 









Net Power 



t S3 






1,8 00 

4 3(., too 







1,8 00 

3: 00 







1,8 00 













10, ISO 


9, 8 so 



1 SO 






1, 800 






3 00 




1 S0.8 1 100.4 1 t2,.8 1 iO,lt8 


9,8:.i i,ioo 

The Following" Is To Be Determined From Steam Chart Or TableSt 
Hi = Heat (Above SaT.) In 1 Lb. Of Steam At Throttle^ A4/:^..B.T.U 
Hji = Heat In 1 Lb. Of Steam At Exhaust (Assuming : 

Adiabatic Expansion)^ /.?/?.. B.T.U 

h^=Heat In 1 Lb. Of Water At Exhaust Temperature 6/...B.T.U. 

Fig. 245. — Showing a form of log sheet for arrangement of steam-turbine test data 
for convenience in computing performance valves. (All of the heat values are above 

hr. From Column 7, Fig. 245, Vr = 9,828 kw. Therefore, by For. (35), 
the water rate, Wr = W/(tPK) = 436,800 ^ (3 X 9,828) = 14.8 lb. per 

(b) The Number Of British Thermal Units Consumed Per Kilo- 
watt-hour may be computed by the following formula: 

Q =Wk (Hi - h2> 

(B.t.u. per kw.-hr.) 


Wherein : Q = the heat, in British thermal units, consumed per kilowatt- 
hour. Wk = the water rate, in pounds of steam per kilowatt-hour. 
Hi = the heat, in British thermal units, in 1 lb. of steam at the throttle. 
h.2 = heat, in British thermal units, in 1 lb. of water at the temperature 
of the exhaust. As stated in Fig. 245, Hi = 1,252 B.t.u., and h2 = 61 
B.t.u. From solution under For. (35), Wk = 14.8 lb. per kw.-hr. There- 
fore, by For. (36), the heat consumed per kilowatt-hour = Q = Wif(Hj — 
hi) = 14.8 X (1,252 - 61) = 17,626 B.t.u. per kw.-hr. 


(c) The Number Of Foot-pounds Of Net Work Developed Per 
British Thermal Unit may be computed by the following formula: 

(37) W = ?J355,000 ^^^ _,^ ^^^ ^^^^ 

Wherein: W = the work, in foot-pounds, developed by 1 B.t.u. Q = 
heat, in British thermal units, consumed per kilowatt-hour. 2,655,000 = 
the mechanical equivalent, in foot-pounds, of 1 kw.-hr. From the solu- 
tion under For. (36), Q = 17,626 B.t.u. per kw.-hr. Therefore, by For. 
(37), the number of foot-pounds of net work developed per British thermal 
unit, W = 2,655,000/Q = 2,655,000 -r- 17,626 = 150 ft.-lb. per B.t.u. 

(d) The Thermal Efficiency Based On Net Generator Output 
(at generator terminals) may be computed by the following formula : 

Q 41 Q 

(38) Er = ^^ (decimal) 

Wherein: Et = the thermal efficiency, exjDressed decimally. Q = the 
heat, in British thermal units, consumed per kilowatt-hour. 3,413 = 
the heat equivalent, in British thermal units, of 1 kw.-hr. From solution 
under For. (36), Q = 17,626 B.t.u. per kw.-hr. Therefore, by For. (38), 
the thermal efficiency, Er = 3,413/Q = 3,413 ^ 17,626 = 0.193, or 19.3 
per cent. 

(e) The Rankine-cycle Ratio May Be Determined by the following 
formula : 

(39) Er = ^,w"^' ~~J^^^ (decimal) 

(Ml — ±±2) 

Wherein: Er = the Rankine-cycle ratio, expressed decimally. Er = 
thermal efficiency, expressed decimally. Hi, H2, and h2 are as specified 
in Fig. 245. From solution under For. (38), Er = 0.193. From Fig. 245, 
Hi, H2, and h2 = 1,252, 894 and 61 B.t.u., respectively. Therefore, by 
For. (39), the Rankine cycle ratio, E^ = [Er(Hi - h2)]/(Hi - H2) = 
[0.193(1,252 - 61)] ^ (1,252 - 894) = 0.193 X 1,191 ^ 358 = 0.642, or 
64.2 per cent. 

Note. — The Computation Of The Performance Values Of A 
Turbine On The Basis Of The Brake Horsepower Output may be 
made in a manner which is substantially the same as that indicated in the 
solution of the above example. The brake horsepower is found by means 
of a brake (see the author's Steam-engine Principles and Practice). 
Then, the value of the brake horsepower or its equivalent is used in the 
following formulas and For. (39). 

(40) w„ = ^ (lb. per hp.-hr.) 

(41) Qh = W^(Hi - ha) (B.t.u. per hp.-hr.) 

(42) ^^M80^0 ( 

(43) Er = -^ (decimal) 


Wherein: W^ = the weight of steam, in pounds, required to develop 
1 hp.-hr. W = the total weight of steam, in pounds, consumed by 
the turbine during the test, t = the duration of the test, in hours. 
Vh = average net power, in horsepower, developed by the turbine. 
Qjj = the heat, in British thermal units, consumed per horsepower-hour. 
W = the work, in foot-pounds, developed by 1 B.t.u. Et = the thermal 
efficiency, expressed decimally. Hi and hi are as specified under Fig. 245. 

241. The Reason Why The Five Different Methods Of 
Expressing The Performance Values Of Steam Turbines 

(Sec. 240) are used in the A.S.M.E. Test Code (Sec. 248) 
is that each method has a somewhat different significance. 
Each is discussed below. No one method has been adopted 
as a standard. Furthermore, various engineers prefer differ- 
ent bases for comparing the performance values of heat 
engines. Also an internal combustion engine does not have a 
water rate or a Rankine-cycle ratio. Hence, methods (2), 
(3) and (4) of Sec. 240 provide the only basis for comparing the 
thermal performance of a steam engine or a steam turbine 
with that of an internal combustion engine. Consequently, 
to provide for every contingency, a complete turbine test 
report should show each of the above mentioned (Sec. 240) 
performance values. See notes below and Sees. 242 to 245. 

Note. — The Water Rate is generally used by turbine manufacturers 
as the basis of their performance guarantees. However, unless the 
initial and final steam conditions are known the water rate is meaningless. 
The reason is that the water rate for a given turbine will vary consider- 
ably with the steam conditions. It is used principally because all of the 
other performance values are determined from it; see Fors. (35) to (39). 
Furthermore, the average turbine purchaser has, through ''handed- 
down" practice, learned to think of steam prime mover economies in 
terms of water rate. Where two turbines operate under the same steam 
conditions, their water rates form an absolute basis for comparison of 
their economies. However, it should be remembered that a low water 
rate does not necessarily indicate a low fuel consumption. 

Note. — The Foot-pounds Per British Thermal Unit And The 
British Thermal Units Per Kilowatt-hour Or Per Brake Horse- 
power-hour are merely different ways of expressing thermal efficiency 
which is discussed in Sec. 245. 

242. The Definitions Of The Terms "Total Heat Input" 
And "Available Heat" should be thoroughly understood 


before one attempts to study the significance of the different 
methods of expressing steam turbine performance values. 
Consequently, these terms are defined and explained in the 
following notes: 

Note. — The Total Heat Input to the turbine per pound of steam 
may be defined as the difference between the heat content, Hi, in British 
thermal units, in 1 lb. of steam at conditions existing at the throttle, and 
the heat content, h2, in British thermal units in 1 lb. of water at the tem- 
perature of the turbine exhaust. That is, 

(44) Total heat in-put -per lb. = (Hi — h2) (B.t.u. per lb.) 
Under the steam conditions tabulated in Fig. 245 (150.8 lb. per sq. in., 
gage, and 100° F. superheat at the throttle, and 92.8° F. at the exhaust), 
the total heat input per pound (Fig. 245) = Hi — h2 = 1,252 — 61 = 
1,191 B.t.u. per lb. (The values of Hi and h2 are taken from steam 
tables.) That is, in considering the total heat input per lb., the tem- 
perature of the exhaust is taken as the starting or datum point. 

Note. — The Available Heat per pound of steam may be defined as 
the difference between the heat content per pound of the steam under 
the steam conditions existing at the throttle. Hi, and the heat content per 
pound of the steam. Ha, after adiabatic expansion to the exhaust pres- 
sure. The amount of the "available" heat per pound of steam may be 
most conveniently obtained by using a steam chart as follows: Find, 
by the chart (Fig. 15) the heat, Hi, in 1 lb. of steam at the initial condi- 
tions. Next, find the heat, H2, in 1 lb. of steam after adiabatic expansion 
to the final condition. The difference between these two values is the 
"available" heat in British thermal units per pound. Expressed as a 
formula : 

(45) The available heat per lb. = Hi — H2 (B.t.u. per lb.) 
Wherein: Hi = the heat, in British thermal units, in 1 lb. of steam at the 
initial steam_ conditions. H2 = the heat, in British thermal units, in 
1 lb. ol steam after it has expanded adiabatically down to the final 
temperature (at the exhaust). Under the steam conditions outlined 
in Fig. 245, values being obtained from the steam chart, Fig. 15, the 
available heat per lb. = Hi - H2 = 1,252 - 894 = 358 B.t.u. per lb. 

The reason this "358 B.t.u. per lb." is called the "available" heat for 
these conditions is because that, with the stated initial and final condi- 
tions, it is all of the heat that is available for conversion into mechanical 
work. It is absolutely all of the heat that could for these conditions be 
converted into work, even in a theoretically perfect or ideal engine. Why 
this is true is explained in the author's Practical Heat. That is, if a 
steam engine could be constructed which was an ideal or theoretically 
perfect engine, it could, under the steam conditions outlined in Fig. 245, 
convert into work only 358 of the 1,252 B.t.u. per lb. which are supplied 
to it; the other (1,252 - 358) = 894 B.t.u. being exhausted. 


243. A Rankine -cycle Efficiency value for a certain set of 
steam conditions indicates the maxiyniun percentage of the 
total heat input (Sec. 242) which a theoretically-perfect ideal 
vapor engine — steam engine or steam turbine — could, when 
operating between these steam conditions, convert into 
mechanical work. That is, 

(46) Rankine-cycle efficiency = 

available heat per lb. 


total heat input per lb. 
or, using symbols; 


(47) Rankine-cycle efficiency = :^ — ~ (decimal) 

Ml — 112 

This efficiency is determined solely by the given steam condi- 
tions. It constitutes an index of the excellence of the steam 
conditions. Certain large electric central station companies 
keep a record of how this efficiency varies from day to day 
and from month to month for their steam prime movers. 
Such a record enables the chief engineers to keep check on — 
and to maintain at maximum effectiveness — the steam con- 
ditions under which the prime movers operate. As indicated 
by For. (46) it is based on the available heat per pound of 
steam (Sec. 242). Note particularly the example below and 
the comments which follow it. 

Example. — What is the Rankine-cycle efficiency for the steam condi- 
tions outlined in Fig. 245? Solution. — By the notes under preceding 
Sec. 242, the available heat for the steam conditions of Fig. 245 is 358 B.t.u. 
per Ih., and the total heat input is 1,191 B.t.u. per lb. Therefore, by 
(For. 46), the Rankine-cycle efficiency = (available heat) /(total heat 
input) = 358 -^ 1,191 = 0.30, or 30 per cent. Note that the values used 
in computing this efficiency are not in any manner dependent upon tho 
operation of the turbine, but only upon the stated initial and final condi- 
tions of the steam. Consequently any old kind of a turbine or engine 
operating under the steam conditions outHned in Fig. 245 would have this 
same Rankine-cycle efficiency of 30 per cent. 

244. A Rankine-cycle Ratio value for a given vapor^ngine — 
steam engine or steam turbine — indicates for the given steam 
conditions, the percentage of the available heat that the g iven 
engine converts into mechanical work.\ It can be determined 
accurately for a given turbine only by testing the turMoe for 


work output and observing simultaneously the supply and 
exhaust steam conditions. ' Expressed as a formula: 

(48) Rankine-cycle ratio — 

Work output in B.t.u. per lb. ., . ,. 

A — T^J-^T — / T, (decimal) 

Available heat per Uh. 

or, using symbols; 

(49) Rankine-cycle ratio = 

Work output in B.t.u. per lb. of steam ,, . ,. 
^ == ^ (decimal) 

Jll — ±±2 

This efficiency is an index of the excellence of design and 
mechanical condition of the turbine. Consequently, a compari- 
son of the Rankine-cycle ratios of different vapor engines 
provides a measure of the excellence of design of the engines 
for the steam conditions under which each is operating and of 
its mechanical condition. Thus even though a turbine be 
excellently designed, if its mechanical condition is permitted 
to deteriorate — if bearings become scored and blading becomes 
clogged or broken — its Rankine-cycle ratio will be low. Con- 
versely, a turbine may be well constructed mechanically and be 
in excellent mechanical condition, but if it is poorly designed 
its Rankine-cycle ratio will be low. 

Explanation. — Consider the turbine, the test results of which are 
tabulated in Fig. 245. Since from the solution under For. (35), 14.8 Ih. 
of steam produce 1 kw.-hr., or 2,655,000 ft.-lb., 1 lb. of steam produces 
(2,655,000 4- 14.8) = 179,392 ft.-lb. Since there are 778 ft.-lb. in 1 
B.t.u., the nutnber of British thermal units which are, from each pound of 
steam, converted into work = 179,392 -^ 778 = 230 B.t.u. per lb. of steam. 
That is, the work output is 230 B.t.u. per lb. of steam. This means that in 
each pound of steam only 230 B.t.u. were actually converted into work; 
whereas (Sec. 242), there were originally, in each pound, 358 B.t.u. which 
were available for conversion into work. By For. (48), (the Rankine- 
cycle ratio) = (Work output in B .t.u.) / (Available heat) = 230 -^ 358 = 
0.642 or 64.2 per cent. This may be explained as follows: If the turbine 
had been "perfectly" designed and was in perfect mechanical condi- 
tion — a theoretically-perfect ideal vapor engine — all of the available 
358 B.t.u. per lb. would have been converted into work. But since the 
turbine only converts 230 B.t.u. per lb. into work, the design and mechan- 
ical condition is only 64.2 per cent, "perfect." 

245. The Thermal Efficiency expresses the percentage of the 
total heat input of the steam consumed by the turbine which is 


converted into work. It is the product of the Rankine-cycle 
efficiency and the Rankine-cycle ratio. Thus, it is a sort of an 
overall efficiency which combines into one value an index of 
the excellence of the heat conditions (Sec. 243) and of the 
excellence of the design and mechanical condition (Sec. 244). 
This combining, into one value, of the expressions for the 
excellence of heat conditions and of design and mechanical 
condition may be understood from the following: 

(50) Thermal eff. = {Rankine-cycle eff.) X 

{Rankine-cycle ratio) (decimal) 
or, using symbols; 

(51) Thermal eff. = 

Hi — Ho Work output in B.t.u. per lb. of steam , , . i\ 

hT^.x hT^^h^ ^^"""^^'^ 

then, simplifying: 

(52) Thermal eff. = 

Work output in B.t.u. per lb. of steam ,, . ,. 
"^ ^ rf ^ (decimal) 

Xll — 112 

It is shown in Sec. 243 how the Rankine-cycle efficiency indi- 
cates the excellence of heat conditions, and in Sec. 244 how the 
Rankine-cycle ratio indicates the excellence of design and 
mechanical condition. Therefore, since the ''thermal effi- 
ciency formula" (For. 51) contains both of these values, it is 
evident that the thermal efficiency value must provide an index 
of the excellence of both heat conditions and design and 
mechanical condition. Hence the heat consumptio7i of turbines 
of different designs may be intelligently compared on the basis 
of their thermal efficiencies even when the turbines are operating 
under different steam conditions.// pThe onewhich has the highest > 
thermal efficiency will require the least heat for its operation- — 
but the one having the highest thermal efficiency may not be^ 
the cheapest to operate because it may cost much more to 
produce a pound of steam for the steam conditions of the 
high-thermal-efficiency turbine than it will for the steam 
conditions of the low-thermal-efficiency turbine ;/see Div. 14. 

Explanation. — Again considering the turbine test results of Fig. 245: 
From Sec. 243, the Rankine-cycle efficiency = 30 per cent. From Sec. 244, 


the Rankine-cycle ratio = 64.2 per cent. By For. (51), the thermal effi- 
ciency = {Rankine-cycle efficiency) X (Rankine-cycle ratio) = 0.30 X 
0.642 = 0.1926, or 19.3 per cent. That is, of the total heat input per 
pound of steam (1,191 B.t.u., Sec. 242), only 30 per cent. (358 B.t.u.) 
could have, by a theoretically-perfect engine, been converted into work. 
Furthermore, this particular turbine (Fig. 245) only converted into work 
64.2 per cent, of the 30 per cent, which could, under ideal conditions, 
possibly have been so converted — or, it converted into work only 19.3 
per cent., of the total heat input. 

246. Graphs Which Show The Total Steam Consumption 
And The Wate r Rate OjLA Turbin e At V^anous Loads {Fig;. 244) 

are very conveniejit^for^comparing (Sec. 249) the operations of 
two^~orlnore turbines; also for comparing test results with the 
manufacturer's guarantee. Such graphs are obtained as 
follows: A complete test of the turbine is made at each of the 
various loads. The total steam consumption and the water 
rate for each of the several loads are computed as in the preced- 
ing example. Then, the total steam consumption, in pounds, 
and the water rate, in pounds of steam per kilowatt-hour or 
per brake horsepower-hour, are plotted (Fig. 244) against the 
load in kilowatts or in brake horsepower. To obtain the 
data for plotting these curves, tests are usually made at each 
of the following percentages of full rated load: 50, 75, 100, 
and sometimes 125 per cent. 

247. In Making A Test On A Steam Turbine It is Desir- 
able That Certain Data Be Taken Whereby Any Operating 
Faults May be Located. — For example, by observing the 
steam pressure in the various stages (Item 10c Sec. 248) 
information may be obtained as to whether or not the blading 
is fouled or whether the diaphragm glands are leaking. Also, 
by comparing the pressure in the exhaust pipe near the turbine 
with that in the condenser, it will be evident whether or not 
the pressure drop in the exhaust pipe is excessive. Ordinarily, 
this pressure drop should not exceed 0.25 to 0.5 lb. per s qT in . 
Other observations which"are not directly essential in determin- 
ing the performance values (Sec. 240) but which may be used 
in locating operating faults are tabulated under the fqirgwing 

248. A Data Form For A Complete Steam Turbine Test is 
embodied in the A.S.M.E. Test Code, which is given below: 



Determine the object of the test (Sec. 226), take the dimensions and 
note the physical conditions not only of the turbine but of the entire plant 
concerned, examine for leakages, install the testing appliances, etc., as 
pointed out in the general instructions given in Pars. 1 to 33 (preceding 
sections of this division) and prepare for the test accordingly. 


The apparatus and instruments required for a performance test of a 
steam turbine or turbo-generator, are: 

(a) Tanks and platform scales for weighing water (or water meters 
calibrated in place). 

(6) Graduated scales attached to the water glasses of the boilers. 

(c) Pressure gages, vacuum gages, and thermometers. 

{d) Steam calorimeter. 

(e) Barometer. 

(J) Tachometer, revolution-counter, or other similar speed-measuring 
apparatus or equipment. 

{g) Friction brake or dynamometer. 

Qi) Voltmeters, ammeters, wattmeters, and watt-hour meters for the 
electrical measurements in the case of a turbo-generator. 

Directions regarding the use and calibration of these appliances are 
given in Pars. 7 to 9, and in Pars. 24 to 33 (A. S. M. E. Test Code, 1915). 
The determination of the heat and steam consumption of a turbine or 
turbo-generator should conform to the same methods as those described 
in the Steam Engine Code, Part V. {See exam-pie under Sec. 240; 
also the author's Steam-engine Principles and Practice.) If the 
steam consump>:ion is determined from the water discharged by the wet 
vacuum or hot-well pump, correction should be made for water drawn 
in through the packing glands of the turbine shaft, for condenser leakage, 
and for any other foreign supply of water. 

The rules pertaining to the subjects Operating Conditions, Duration, 
Starting and Stopping, Records, and Calculation of Results, are identically 
the same as those given under the respective headings in the {A. S. M. E.) 
Steam Engine Code, Pars. 71 to 77, with the single exception of the 
matter relating to indicator diagrams and results computed therefrom; 
and reference may be made to that code for the directions required in 
these particulars. 


The data and results should be reported in accordance with the form 
(Table 11) given herewith, adding lines for data not provided for, or 
omitting those not required, as may conform to the object (Sec. 226) 
in view. If a shorter form of report is desired, the items in fine print 
designated by letters of the alphabet, may be omitted; or if only the prin- 


cipal data and results are desired, the subjoined abbreviated table 
(Table 12) may be used. Unless otherwise indicated, the items should 
be the averages of the data. 

Table 11. Data And Results Of Steam Turbine Or 
Turbo -generator Test 

Code of 1915 

1. Test of turbine located at 

To determine 

Test conducted by 

2. Type of turbine (impulse, reaction, or combination) 

(a) Number of stages 

(6) Condensing or non-condensing 

(c) Diameter of rotors 

{d) Number and type of nozzles 

(e) Area of nozzles 

(/) Type of governor 

3. Class of service (electric, pumping, compressor, etc.) 

4. Auxiliaries (steam or electric driven) 

(a) Type and make of condensing equipment 

(6) Rated capacity of condensing equipment 

(c) Type of oil pumps (direct or independently driven) 

(d) Type of exciter (direct or independently driven) 

(e) Type of ventilating fan, if separately driven 

5. Rated capacity of turbine 

(a) Name of builders 

6. Capacity of generator or other apparatus consuming power of 


Date And Duration 

7. Date 

8. Duration hr. 

Average Pressures And Temperatures 

9. Pressure in steam pipe near throttle by gage lb. per sq. in. 

10. Barometric pressure in. of mercury 

(o) Pressure at boiler by gage lb. per sq. in. 

(6) Pressure in steam chest by gage lb. per sq. in. 

(c) Pressure in various stages lb. per sq. in. 

11. Pressure in exhaust pipe near turbine, by gage lb. per sq. in. 

12. Vacuum in condenser in. of mercury 

(o) Corresponding absolute pressure lb. per sq. in. 

(6) Absolute pressure in exhaust chamber of turbine lb. per sq. in. 

13. Temperature of steam near throttle deg. 

(a) Temperature of saturated steam at throttle pressure deg. 

(6) Temperature of steam in various stages, if superheated deg. 

14. Temperature of steam in exhaust pipe near turbine deg. 

(a) Temperature of circulating water entering condenser deg, 

(6) Temperature of circulating water leaving condenser deg. 

(c) Temperature of air in tuybine room deg- 

18 -^/^ ^'" ' '"'*■ i,-- 


\^^/ly ni\/ic\OH 


Quality Of Steam 

15. Percentage of moisture in steam near throttle, or number of degrees 

of superheating per cent, or deg. 

16. Total water fed to boilers lb. 

17. Total condensate from surface condenser (corrected for condenser 

leakage and leakage of shaft and pump glands) lb. 

18. Total dry steam consumed (Item 16 or 17 less moisture in steam) 


Hourly Quantities 

19. Total water fed to boilers or drawn from surface condenser per 

hour lb. 

20. Total dry steam consumed for all purposes per hour (Item 18 -^ 

Item 8) lb. 

21. Steam consumed per hour for all purposes foreign to the turbine 

(including drips and leakage of plant) lb. 

22. Dry steam consumed by turbine per hour (Item 20 — Item 21) ... . 


(o) Circulating water supplied to condenser per hour lb. 

Hourly Heat Data 

23. Heat units consumed by turbine per hour [Item 22 X (total heat 

of steam per pound at pressure of Item 9 less heat in 1 lb. of 
water at temperature of Item 14)] B.t.u. 

(a) Heat converted into work per hour B.t.u. 

(6) Heat rejected to condenser per hour [Item 22a X 

( Item 146 — Item 14a)] (approximate) B.t.u 

(c) Heat rejected in the form of steam withdrawn from the turbine. . . .B.t.u 

(d) Heat lost by radiation from turbine, and unaccounted for B.t.u 

Electrical Data 

24. Average volts, each phase volts 

25. Average amperes, each phase amperes 

26. Average kilowatts, first meter kw. 

27. Average kilowatts, second meter kw. 

28. Total kilowatts output , kw. 

29. Power factor 

30. Kilowatts used for excitation, and for separately driven ventilating 

fan kw. 

31. Net kilowatt output kw. 


32. Revolutions per minute r.p.m. 

33. Variation of speed between no load and full load r.p.m. 

34. Momentary fluctuation of speed on suddenly changing from full 

k>ad to half-load r.p.m. 



35. Brake horsepower, if determined b.hp. 

36. Electrical horsepower e.hp. 

Economy Results 

37. Dry steam consumed by turbine per b,hp.-hr lb. 

38. Dry steam consumed per net kw.-hr lb. 

39. Heat units consumed by turbine per b.hp.-hr. (Item 23 -h Item 35) 


40. Heat units consumed per net kw.-hr B.t.u. 

Efficiency Results 

41. Thermal efficiency of turbine (2,546.5 ^ Item 39) X 100 

per cent. 

42. Efficiency of Rankine cycle between temperatures of Items 13 and 

14 per cent. 

43. Rankine cycle ratio (Item 41 -r- Item 42) 

Work Done Per Heat Unit 

44. Net work per B.t.u. consumed by turbine (1,980,000 -r- Item 39). . . 

Table 12. Principal Data And Results Of Turbine Test 

1. Dimensions 

2. Date 

3. Duration hr. 

4. Pressure in steam pipe near throttle by gage lb. per sq. in. 

5. Vacuum in condenser in. of mercury 

6. Percentage of moisture in steam near throttle or number of degrees 

of superheating per cent, or deg. 

7. Net steam consumed per hour lb. 

8. Revolutions per minute r.p.m. 

9. Brake horsepower developed b.hp. 

10. Kw, output kw. 

11. Steam consumed per b.hp.-hr lb. 

12. Heat consumed per b.hp.-hr B.t.u. 

13. Steam consumed per kw.-hr lb. 

14. Heat consumed per kw.-hr B.t.u. 

249. A Comparison Of The Performances Of Different 
Steam Turbines, or of the same turbine at different times, 
cannot be intelligently made if the computations of the 
performance values are based on different steam conditions, 
such as different initial pressures and temperatures, and differ- 


ent exhaust pressures and temperatures. Usually, it is 
impractical to make two tests of the same turbine, or tests of 
different turbines under the same steam conditions. Conse- 
quently, to make a fair comparison between two or more 
sets of performance values, it is usually necessary to apply 
certain corrections. Such corrections should be applied 
which will convert the performance values which are made 
under one set of steam conditions to those which would obtain 
under some other set of steam conditions. The amount of 
the corrections and the method of their application are 
treated in Div. 13. 


1. What is the purpose of testing a steam turbine? 

2. For what purposes may the performance values as computed from the results of a 
turbine test be used? 

3. What should govern the conditions under which a test is made? If the object of 
the test is to determine how nearly the actual operating performance complies with the 
guaranteed performance, what are the conditions which should obtain? 

4. Name five of the more important data items which should be observed in testing 
a steam turbine. 

5. Over how long a period of time should a turbine test be extended? At what time 
intervals should the instrument readings be noted and recorded? 

6. Why should all instruments used in a turbine test be calibrated both before and 
after the test? 

7. What properties determine the condition of the steam entering the turbine? 

8. Explain how the properties which determine the condition of the steam entering 
the turbine are measured. 

9. What property of the steam at the exhaust must be known? 

10. Why is it generally desirable to determine the pressure at the exhaust flange of the 

11. What is meant by "referred to a 30-in. barometer?" 

12. Name two methods of determining the power output of a turbine. 

13. What instruments may be used to determine the power output of a turbine 
electrically? Which instruments are preferable if the load remains constant? If the 
load fluctuates? 

14. In testing a turbo-alternator, why is it desirable that the power factor be unity? 
What kind of a load will give a unity power factor? 

16. How is the net power output in watts of a turbo-alternator determined if the 
exciter is mounted on the generator shaft? If the alternator is excited from a separately 
driven exciter? 

16. Name three methods of determining the quantity of steam consumed by a turbine. 

17. Make a sketch of the apparatus required for determining the steam consumption 
of a turbine which is operated in conjunction with a surface condenser by weighing the 

18. Explain how the condenser leakage may be determined. If the condenser leakage 
is excessive, what should be done before proceeding with the test of the turbine? 

19. Make a sketch of the apparatus for determining the steam consumption of a turbine 
by weighing the boiler feed water. 

20. In determining the steam consumption of a turbine by weighing the boiler feed 
water, what leakages must be determined? Explain how the amount of this leakage may 
be measured. 


21. Why is it usually undesirable to use a steam flow meter to determine the total I 
steam consumption of a turbine? 

22. How may the quantity of the condensate of a jet condenser be measured? 

23. How is the speed of rotation of a turbine rotor determined? 

24. In what five ways are steam turbine performance values frequently expressed? ' 

25. Explain how each is computed from the test data. 

26. Why are the five different methods of expressing a turbine performance included ^ 
in a test report? J 

27. Why is the water rate used to express turbine performance? ; 

28. For what purpose may the ft. -lb. per B.t.u. and the B.t.u. per kw.-hr. be used? 

29. Define the terms, total heat input and available heat. 

30. What is indicated by a Rankine cycle efficiency value? i 

31. What is indicated by a Rankine cycle ratio value? To what is it an index? ' 

32. What is indicated by the thermal efficiency? To what is it an index? | 

33. For what purposes may graphs, which show the total steam consumption and the | 
water rate of a turbine at various loads, be conveniently used? Explain how such graphs | 
may be obtained. \ 

34. If two or more turbines have been tested under different steam conditions, what j 
must be done before their performance values can be intelligently compared? ! 

35. Make a sketch showing location of all instruments used in testing a turbo-alter- \ 
nator which is operated in conjunction with a surface condenser. i 



250. The Water Rate And Thermal Efficiency Of A Turbine 
Are Dependent On The Conditions Of The Supply And 
Exhaust Steam. — In general, it may be said that the greater 
is the heat content of the supplied steam and the smaller is 
the heat content of the exhaust steam, the higher will be the 
thermal efficiency of the turbine and the lower will be its 
water rate. Hence, those factors which produce great heat 
content in steam — high pressure, high quality, and high 
superheat — are to be desired as properties of the supply 
steam. Also, those factors which produce small heat content 
in the exhaust steam — low exhaust pressure (high vacuum) 
and little steam friction and leakage within the turbine — are 
very desirable. Unfortunately, however, it always costs 
more to produce supply steam of great heat content — high 
pressure and superheat — than it does to produce supply 
steam of small heat content. Likewise, the condensers, 
cooling water, and auxiliary power for high-vacuum service 
cost more than for low-vacuum service. Hence, it is the 
object of this division to study the several effects of the above 
steam conditions on the efficiency of turbines and on their 
cost of operation so that the most economical conditions for 
any given turbine may be determined. Figs. 245A and 2455 
illustrate the steam conditions in a large turbine. 

Note. — The Effects Of Pressure, Superheat, And Vacuum On 
The Water Rate And Thermal Efficiency Of A Theoretically 
Perfect Turbine will first be discussed because the effects in a theoretic- 
ally perfect turbine are explanatory of the effects in an actual or com- 
mercial turbine. Wherever the effects in an actual turbine are different 
from those in the theoretical, these differences will be explained at a 
later point in this text. 

Explanation. — The water rate of a theoretically perfect turbine is 
given in For. (19) which is restated below as For. (53). The thermal 
eflBciency of a theoretically perfect turbine, which is the same as its 




Rankine cijcle efficiency, is given by For. (54) which is derived in the 
author's Practical Heat. 

(53) W H = ^' Z^ (lb. per hp.-hr.) 



Hi — H2 


Hi - ha 

Wherein : Wh = the turbine water rate, in pounds per horsepower-hour. 
Er = the turbine's thermal efficiency, expressed decimally. Hi = the 
total heat of 1 lb. of supply steam, in British thermal units. H2 = the 
total of 1 lb. of steam after adiabatic expansion to the exhaust pressure, 
in British thermal units, ha = the heat of 1 lb. of water at the tempera- 
ture which is the boiling point at the exhaust pressure, in British thermal 
units. Hi and H2 may, as explained in Div. i, be found from a total- 
heat-entropy chart (Fig. 15); ha is found from the steam tables. 

Inspection of For. (53) shows that the greater is the difference between 
Hi and Ho, for a given turbine, the smaller will be the water rate of the 
turbine. Hence, changes in the steam conditions which increase Hi or 
which decrease H2, will enable the turbine to operate with a lower water 
rate — and vice-versa. With regard to For. (54), however, it is not evi- 
dent by inspection just what effects on the thermal efficiency will be 
produced by changes in the steam conditions. To illustrate the effects 
of changing the quality, pressure, and superheat of the supply steam 
and of changing the exhaust pressure (vacuum), the specific examples 
following Table 251 are here given. 

251. Table Showing The Effect Of Different Steam Condi- 
tions On The Water Rate And Thermal Efficiency Of A 
Theoretically Perfect Steam Turbine. — The method of com- 
puting these values is shown in the following examples. 

Supplied steam 

Exhaust steam 


Thermal or Rankine- 
cycle efficiency, per 

lb. per 
SCJ. -in. 

































































Fig. 245A. — Steam conditions in a 40,000 kw. Westinghouse turbine when it is 
delivering 28,000 kw. The primary valve is admitting steam at 250 lb. per sq. in. The 
secondary valve is just beginning to open, The tertiary valve is closed. {Power, 
August 8, 1922.) 



Fig. 24oJB. — Continuation of Fig. 245J.. 


Example. — First Condition.' — Supplied steam pressure, 150 Ih. per 
sq. in. gage; quality, 90 per cent.; vacuum, 28-in. mercury column. What 
are the water rate and the efficiency of this perfect turbine? Solution. — 
From the total-heat-entropy chart of Fig. 15, the value of Hi is found 
at the intersection of the 90 per cent, quahty Hne and the 150-lb. pressure 
line to be 1,100 B.t.u. Following (on Fig. 15) vertically downward to 
the 28-in. vacuum line, H2 is found to be 813 B.t.u. From a steam table, 
hi is found to be 94 B.t.u. Hence, by For. (53): the water rate = W^ = 
2,545/(Hi - H2) = 2,545 ^ (1,100 - 813) = 8.58 lb. per hp.-hr. This 
result could also have been read from the scale A at the top of Fig. 15. 
Also, by For. (54): The thermal efficiency = Et = (Hi — H2)/(Hi — ha) = 
(1,100 - 813) -^ (1,100 - 94) = 0.292 or 29.2 percent. 

Example. — Second Condition. — Supplied steam pressure, 150 lb. per 
sq. in. gage; quality, dry saturated; vacuum, 28-in. mercury column. What 
are the water rate and thermal efficiency of the turbine under these con- 
ditions and how much have they been improved? Solution. — In the 
same manner as in the first condition, the water rate is found to be 7.84 lb. 
per hp.-hr., and the thermal efficiency to be 29.5 per cerit. Hence, the 
decrease in water rate = 8.58 - 7.84 = 0.74 lb. or (0.74 ^ 8.58) = 0.086 
or 8.6 per cent. Also, the increase in efficiency = 29.5 — 29.2 =0.3 per 
cent., or an improvement of (0.3 -r- 29.2) = 0.01 = 1 per cent. 

Example. — Third Condition. — Supplied steam pressure, 175 lb. per 
sq. in. gage; quality, dry saturated; vacuum, 28-in. mercury column. What 
are the water rate and thermal efficiency of the turbine under these con- 
ditions and how much have they been improved? Solution. — In the 
same manner as for the first condition, the water rate is found to be 7.60 
lb. per hp.-hr., and the thermal efficiency to be 30.3 per cent. Hence, the 
decrease in water = 7.84 - 7.60 = 0.24 lb. or (0.24 ^ 7.84) = 0.03 or 3 
per cent. Also, the increase in efficiency = 30.3 — 29.5 =0.8 per cent., 
or an improvement of (0.8 -h 29.5) = 0.027 or 2.7 per cent. 

Example. — Fourth Condition. — Supplied steam pressure, 175 lb. per 
sq. in. gage; superheat, 150° F.; vacuum, 28-in. mercury column. What are 
the water rate and thermal efficiency of the turbine under these conditions 
and how much have they been improved? Solution. — In the same 
manner as for the first condition, the tvater rate is found to be 6.98 lb. per 
hp.-hr., and the thermal efficiency to be 30.7 per cent. Hence, the decrease 
in water rate = 7.60 - 6.98 = 0.62 lb. .or (0.62 ^ 7.60) = 0.081 or 8.1 
per cent. Also, the increase in efficiency — 30.7 — 30.3 =0.4 per cent., 
or an improvement of (0.4 -i- 30.3) = 0.013 or 1.3 per cent. 

Example. — Fifth Condition. — Supplied steam pressure, 175 lb. per 
sq. in. gage; superheat, 150 °F.; vacuum, 29-in. mercury column. What 
are the water rate and thermal efficiency of the turbine under these con- 
ditions and how much have they been improved? Solution. — In the 
same manner as for the first conditions, the water rate is found to be 6.38 
lb. per hp.-hr., and the thermal efficiency to be 32.9 per cent. Hence, the 
decrease in water rate = 6.98 - 6.38 = 0.60 lb. or (0.60 H- 6.98) = 0.086 
or 8.6 per cent. Also, the increase in efficiency = 32.9 — 30.7 = 2.2 per 
cent., or an improvement of (2.2 -4- 30.7) = 0.071 or 7.1 per cent. 


252. Theoretically, The Water Rate And Thermal Effi- 
ciency Of A Turbine Depend Only On The State Of The 
Supply Steam And On The Exhaust Pressure Or Vacuum. — 
How the initial steam pressure and quality or superheat 
and the vacuum affect the water rate and efficiency is shown 
by the preceding typical examples. It is to be noted from the 
examples and from Table 251 that the increase in efficiency 
with changed conditions is not of the same magnitude as is 
the decrease in water rate. These examples show clearly that 
the water rate alone should not be taken as a measure of a 
turbine's thermal performance — as a measure of the fuel that 
must be consumed to insure its operation. 

253. Actually, The Water Rate And Thermal Efficiency 
Depend Also On The Amount Of The Losses Within The 
Turbine. — As stated in Sec. 15, losses occur within a turbine 
casing due to several causes, the principal ones being steam 
friction, steam leakage, eddy currents, radiation, and the 
velocity of the exhaust steam. All of these losses except that 
due to radiation tend to increase the value of H2 in Fors. 
(53) and (54) ; hence, they tend to increase the water rate and 
decrease the efficiency. Furthermore, all of these losses are 
dependent on the quality of the steam in the various passages 
of the turbine (as explained further hereinafter) — the drier 
the steam, the less are the losses. Now, in any turbine, the 
quality of the steam decreases rapidly as the steam flows through 
its passages. Hence, any change, so made in the condition 
of the supply steam as to increase the quahty of the steam in 
the turbine passages, is certain to reduce the losses within the 
turbine and to thereby decrease its water rate and increase 
its thermal efficiency. 

Note. — The Percentage Losses Are Greater In Turbines Of 
Small Than In Those Of Large Capacity. — That this is true is shown 
by the variation of the efficiency ratio, E^, in Fig. 20. The explanation 
of the variation in losses lies in the fact that the interior-surface areas of 
a turbine and the places of possible leakage are greater (in proportion to 
the amount of steam used by the turbine) in small-capacity turbines 
than in large-capacity turbines. 

254. Every Turbine Is Designed For Specific Steam 
Conditions and will perform most efficiently when operated 


under those conditions. The actual conditions under which a 
turbine will operate most efficiently may or may not be the 
same as those conditions for which the turbine was furnished 
by its manufacturer and which are stamped on its name plate ; 
this is because some shop standardization is necessary in 
turbine building and each turbine cannot be specially 
designed for the purchaser. It follows that, in general, a tur- 
bine should always be operated under the steam conditions for 
which it was designed. Hence, the efficiency of turbines will 
not always be increased by increasing the supplied steam pres- 
sure, superheat, or the vacuum. In fact, if too great a 
departure is made from the conditions for which the turbine is 
designed, the efficiency may be decreased instead of increased, 
as explained below. Hence, the manufacturer of a turbine 
should always be consulted as to the effects of condition changes 
before any material changes are made. The manufacturer can 
advise definitely as to whether or not your contemplated 
change is feasible and also as to the economies which will 
thereby be effected. 

Explanation. — Why A Turbine Should Be Operated Only Under 
The Steam Conditions For Which It Was Designed may be explained 
thus : Any change in the steam conditions will, as shown below, increase 
the losses in the turbine. If the steam pressure at the throttle is increased 
and the amount of the superheat and vacuum are unchanged, or if the 
vacuum is increased and the amount of superheat and pressure are un- 
changed, the pressure range of the turbine, or the pressure drop through 
it, is increased. Consequently the pressure drop in each stage is increased 
causing the steam to strike the blades with a greater velocity than that 
for which they were designed. Any change in the value of this velocity 
causes the steam to hit the blades at an angle instead of tangentially 
thereby increasing the loss due to impact. There is also a loss due to the 
increase in the amount of moisture in the steam near the exhaust but 
this loss also occurs in a turbine designed for the improved conditions. 
Increasing the vacuum has the further disadvantage of increasing the 
volume of the exhaust steam. This means that the velocity of the steam 
in the passages near the exhaust end of the turbine must be increased, and 
produces a loss due to exit velocity and to the increased friction. 

An increase in the amount of superheat, with the amount of the 
vacuum and pressure unchanged, increases the volume of the steam 
that must pass through the turbine per unit of time. The only manner 
in which this greater volume of steam can be forced through the passages 
is by increased velocity. A greater velocity means larger friction and 



impact losses. The capacity of the turbine may even be reduced if the 
amount of superheat is increased too much. The resulting losses may 
then offset, to a greater or less degree, the increase in efficiency which the 
improved steam conditions should theoretically provide. 

255. The Capacity Of Any Existing Turbine May Be 
Increased By Increasing The Supply Pressure, Superheat, 
And Vacuum — any one, two or all three. But while the 
capacity of the machine will be thereby increased, it will 
usually be at the expense of efficiency. Just what will be the 
effect on economy of such an increase in capacity is determined 
by the design and construction of the turbine. The manu- 
facturer can furnish exact information. 

256. Table Showing Factors For Computing The Approxi- 
mate Change In The Water Rate Of A Turbme With Changed 
Steam-supply Pressure, Superheat, And Vacuum. — The appli- 
cation of these factors is explained and illustrated in following 

Change in steam condition 

Change in water rate 

Supply-steam Pressure. 
(Increasing the supply-steam pressure 
decreases the water rate and vice 


Turbines up to 1,000 kw. — 1.5 per cent, for 
each 10 lb. per sq. in. change in pressure 

Turbines over 1,000 kw. — 1.0 per cent, for 
each 10 lb. per sq. in. change in pressure. 


(Increasing the superheat decreases the 
water rate and vice versa.) 


Back pressure. 
(Increasing back pressure 

increases the water rate 

and vice versa.) 


(Increasing the vacuum 

decreases the water rate 

and vice versa.) 

Up to 100° F. superheat — 1.0 per cent, for each 
10° F. of change in superheat. 

100° to 150° F. superheat — 0.8 per cent, for 
each 10° F. of change in superheat. 

150° F. to 250° F. superheat — 0.6 per cent, for 
each 10° F. of change in superheat. 

Up to 15 lb. per sq. in. gage — 2 to 3.5 per cent, 
for each pound of back-pressure increase 
(see Fig. 251). 

Between 25 and 27 in. 
of vacuum. 

5 per cent, per inch 

Between 27 and 28 in. — 6 per cent, per inch 
of vacuum. 

Between 28 and 29 in. — 10 per cent, per inch 
of vacuum. 


sections. The preceding factors are approximately correct 
for condition changes within reasonable limits, whether or not 
the tm'bine is changed to suit the new conditions; see note 

Note. — Exact Values Indicating The Effects Of Changing 
Steam Conditions cannot be given because the exact values depend on 
the design and construction of the turbine under consideration and upon 
the steam conditions — pressure, superheat, and vacuum — prior to chang- 
ing the steam conditions. For any given turbine, exact factors, in the 
form of graphs similar to Figs. 252, 253, and 254, may be obtained from 
the manufacturer. 

257. Turbines Are More Efficient When, Other Things 
Being Equal, They Are SuppUed With Steam At High Pres- 
sure. — As suggested by Fig. 
246 and also by Fig. 15, the 
higher the pressure of steam, 
the more heat per pound there 
is in it. That is, the higher 
the pressure, the greater will 
be the value of Hi in Fors. 
(53) and (54). The greater 
the value of Hi — other things 
being equal — the smaller will 
be the water rate and (gener- 
ally) the greater will be the 
thermal efficiency. This is 
brought out by Table 251 and 
by the examples which follow 
it. But if the turbine is to 
efficiently use the high-pres- 
sure steam, it must have been 

designed (Sec. 254) for that pressure. 

258. The Effect Of Increasing The Supply-steam Pressure 
Of An Existing Turbine is generally to increase, to some extent, 
the efficiency of the turbine. But, the turbine may require 
new nozzles for the higher pressure and, if the turbine is 
already operating on steam at a pressure near that for which 
it is designed, or if the turbine is operated most of the time 
at fractional loads, the efficiency may be increased but slightly 




































Temperature, Deg.Fahr 
Fig. 246. — Graph showing the varia- 
tion of the temperature of saturated 
steam with the steam pressure. 


or it even may be decreased by increasing the supply pressure. 
Furthermore, steam at higher pressures costs more to produce 
than does steam at lower pressures — the boiler losses are 
greater, and more expensive boilers must be used. Hence, 
to determine whether a change to a considerably higher 
steam-pressure is advisable, it is best to consult the turbine 
builder's engineering department. 

Note. — The Steam Pressures Which Are Advisable For Turbine 
Operation are as follows : For small turbines, say up to 200-kw. capacity, 
about 150 to 175 lb. per sq. in. gage. For medium-capacity turbines, 
say 200 to 5,000 kw., about 200 to 250 lb. per sq. in. gage. For large 
capacity turbines, as in the large central stations the tendency is con- 
tinually toward higher pressures — some now use pressures as high as 
350 lb. per sq. in. gage. It is doubtful whether pressures higher than 
400 lb. will be used, however, because of the high cost and maintenance 
expense of boilers for these high pressures and because the thermal gains 
from further pressure increase are very small. Note from Fig. 246 that 
the steam temperature — ^which determines, somewhat, the value of Hi 
in For. (53) — increases very slowly with the pressure for pressures 
exceeding 400 lb, per sq. in. 

259. To Compute The Effept On A Turbine's Water Rate Of 
Changing Its Supply Pressure, the factors given in Table 
256 may be used whenever manufacturers' correction curves 
(Sec. 268) are not obtainable. The factors in Table 256 are 
to be used only for computing the effect of changes which do 
not exceed 10 to 15 per cent, of the rated steam pressure. 

Example. — The rated steam-supply pressure for a 2,000-kw. turbo- 
generator is 175 lb. per sq. in. gage (the superheat and vacuum may, 
within reason, be any whatsoever). The water rate of the machine is 
17 lb. per kw.-hr. What water rate may be expected if the steam pres- 
sure is raised from 175 to 200 lb. per sq. in. gage? Solution. — The 
increase in pressure is: 200-175 = 25 lb. per sq. in. Now, 25 4- 10 = 2.5. 
Since, from Table 256, a 1 per cent, decrease in water rate may be ex- 
pected for each 10 lb. per sq. in. pressure increase, the decrease in this 
case will be 2.5 per cent. YiQjiQQ, VaQ luater-rate decrease = 0.025 X 17 = 
0.43 lb. per kw.-hr. Therefore, at 200 lb. per sq. in. pressure, the water 
rate = 17 - 0.43 = 16.57 lb. per kw.-hr. 

260. Turbines Are More Efficient When, Other Things 
Being Equal, They Are Supplied With Highly Superheated 
Steam. — For a given pressure, the value of Hi, Fors. (53) and 


(54), increases with the superheat. Hence the water rate 
decreases with the superheat and (usually) the thermal efficiency 
increases. In using high superheat, however, one must be 
careful that the superheat is not so high that it causes the 
exhaust steam to be superheated — this would result in a 
loss. In general, it may be said that high pressures with 
moderate superheat are more economical than moderate 
pressures with high superheats. 





2 3 A 









68 46 28 











1501 160 1120 175 130 

6 1 2 1 4 1 6.5 1 8 1 10.5 


Fig. 246^. — Showing the condition of the steam in each stage of a ten-stage turbine. 
{General Electric Review, March, 1918.) 

Steam turbines are, inherently, exceedingly well adapted to 
the economic use of superheated steam. 

Note. — The Reason Why Superheating Its Supply Steam Im- 
proves The Economy Of A Turbine is that the superheating reduces 
the water-vapor friction in the turbine: As steam expands in passing 
through successive stages in a turbine and gives up heat which makes the 
rotor turn, the quaHty of the steam tends to become reduced (Fig. 246^1). 
That is, the steam tends to condense and produce water vapor — minute 
drops of water in suspension. The friction of the turbine's rotating disks 
in such a dense wet vapor is considerably higher than in dry steam. 
Similarly the friction of *'wet" steam passing through the nozzles and 
buckets is greater than that of dry steam. This friction represents 
wasted energy. The higher the quality of the steam the less the friction. 
The more the supply steam is superheated the further it will travel 



through (the more heat it can give up in) a given turbine without con- 
densation — without its becoming saturated. Hence, even a little super- 
heating, of the supply steam for a turbine is very valuable. Also, super- 
heating has the added advantage of minimizing blade and nozzle erosion 
in a turbine. 

261. The Effect Of Increasing The Supply-steam Superheat 
Of An Existing Turbine is generally to increase, to some extent, 
the efficiency of the turbine. The principal effect of increasing 
the superheat is to decrease the amount of moisture (water) 
in the steam in the several passages of the turbine (see preceding 
note) ; hence, superheating decreases the amount of the losses 
within the turbine. The principal objection to the use of highly 


10 20 1)0 40 50 60 70 60 90 100 110 120 ITiO 140 150 160 170 160 190 200> 
Superhecit, Degrees Fothrenhelt 

247. — Graph showing the effect of superheat on steam consumption of non-con- 
densing steam turbines. {Sturtevant Company.) 

superheated steam is that, especially in some types of turbines 
(those which have many rows of blades), the high-pressure 
end of the turbine becomes heated to such a high temperature 
that the casing is severely strained. Because turbines have 
no rubbing surfaces which are exposed to the high-pressure 
steam (as have steam engines), there are no lubrication diffi- 
culties occasioned by the use of superheated steam in turbines. 
In any case, however, the cost of superheating the steam (see 
Fig. 248, which is explained hereinafter) must be balanced 
against the gain in efficiency which is produced. The net 
economic value of superheating is thus determined. 

Note. — The Superheats Which Are Most Advisable For Tur- 
bine Operation may roughly be taken as two-thirds of the steam-supply 
pressure in pounds per square inch gage. Thus, about 125° to 150° F. of 


superheat is advisable for medium-sized plants whereas superheats of 
about 200° F. are used in large central stations. Furthermore, non-con- 
densing turbines generally require more superheat and are benefited more 
thereby than condensing turbines. However, unusual local conditions 
such as very-high or very-low fuel cost may render the above values 
inapplicable. Each case should be considered individually on its merits. 
The effect of superheat on a non-condensing turbine is shown in Fig. 247. 

'0^ 20 40 60 80 100 120 140 160 180 200 220 240 
Superheat, Degrees f. 

Fig. 248. — Showing typical relation of power-production cost to superheat. This 
graph is plotted for certain conditions (175 lb. per sq. in. pressure and 210° F. feed water 
in a certain plant) but the general principle which it illustrates is characteristic. For 
these conditions, the greatest decrease in net cost at F, due to superheating occurs with 
a superheat of 160° F. The net decrease in cost, EF = {Decrease in fuel and water cost, 
EG) — {The increase fixed charge and maintenance cost, ED). That is, to determine the 
locations of the points along in OB, for each different superheat, the corresponding 
vertical distance between OC and OH is laid off vertically downward, that is subtracted, 
from OA. 

262. The Actual Net Fuel Saving Due To Superheating A 
Turbine's Supply Steam is usually about 2 to 5 per cent, per 
100° F. of superheat increase, the superheating to be within 
practical limits. Excessive superheating is not economical 
(Fig. 248) because the increased cost of the fuel required and 
the additional expense of equipment for producing and trans- 
mitting the superheated steam, more than offsets the decreased 


fuel consumption due to its use. Advisable superheats are 
given in the preceding note. 

Example. — In the plant and for the conditions for which Fig. 248 was 
plotted, the most economical superheat (at F) is 160° F. With this super- 
heat the net cost of power production is 4 per cent, less than if no super- 
heat were employed. With less superheat than 160° F., as shown by FO, 
or with more superheat than 160° F., as shown by FB, the net cost of 
power is greater. 

263. To Compute The Effect On A Turbine's Water Rate 
Of Changing The Superheat, the factors given in Table 256 
may be used whenever manufacturer's correction curves 
(Sec. 268) are not obtainable. The method of computing the 
effects of superheat changes is illustrated by the following 

Example. — A certain turbine (the supply-steam pressure and the vac- 
uum may be any within reason) shows a water rate at full load of 14 lb. 
per hp.-hr. when supplied with steam of 50° F. superheat. What would 
be its water rate if the superheat were raised to 150° F.? Solution. — 
By Table 256 each 10° F. of superheat increase between 0°F. and 100° F. 
decreases the water rate by 1 per cent., and each 10° F. of superheat 
increase between 100° F. and 150° F. decreases the water rate by 0.8 per 
cent. Hence, for this turbine, the percentage decrease in water rate = 
[(100 - 50) X 1] ^ 10+ [(150 - 100) X 0.8] ^ 10 = 5 + 4 = 9 per cent. 
Hence, the pounds decrease in water rate = 0.09 X 14 = 1.23 lb. per hp.-hr. 
Therefore, the water rate with 150° F. superheat = 14 - 1.26 = 12.74 lb. 
per hp.-hr. 

264. High Vacuum Is The Most Essential Requirement 
For Economical Steam-turbine Operation; see Table 251 and 
the examples which follow it. Maintaining a high vacuum 
provides the most effective method of insuring good economy 
of condensing turbines. Condensing turbines are, in general, 
more economical — often much more; see Div. 14 — than are 
condensing reciprocating engines, principally because the 
turbine is inherently better adapted to the useof high vacuums; 
see below. As a general rule, it pays to keep the vacuum in a 
turbine's exhaust pipe at as high a value as the plant conditions 
and water supply will permit. However, it may not always 
pay to circulate all the water which the condenser pumps can 


handle — above a certain vacuum the cost of pumping addi- 
tional water may be greater than the fuel saving due thereto. 
This is particularly true when the turbine is operating under 
partial load, or in winter when the circulating water is very cold. 
Here again, each turbine is deserving of a separate economic 
study to determine the most economical vacuum in different 

Note. — Turbines Can Effectively Utilize Higher Vacuums Than 
Engines For Two Reasons: (1) Turbine parts are always subjected to 
steam at the same pressure; the low temperatures of the exhaust pressure 
cannot reach back into the hotter parts of the machine whereas in steam 
engines the cylinders are exposed alternately to wide differences of tem- 
perature — this causes cylinder condensation. (2) The steam expansion 
is not limited in the turbine whereas, in the engine, expansion is limited by 







7 Curve 















1 • 






1 '''^ 



vk Line 



Vacuum Llnes^--^ 





YT-^ ■ ^ . \ 1 

^' Zero Pressure Absolute' 

1 1 1 1 1 1 









Fig. 249. — Pressure-volume curves for steam engines and turbines. 

the cylinder volume. (If, with a steam engine, an effort is made to pro- 
vide very-great cylinder volume, the low-pressure cylinder will then 
become excessively large. The friction and other losses which the very 
large cylinder would introduce, much more than offset the economies 
which would theoretically result from the increased cylinder volume.) 
Take, as an example, a condensing steam engine which has an ideal dia- 
gram as shown in Fig. 249 at ABCDG. Since the expansion is limited 
by the cylinder volume, a higher vacuum will increase the diagram area 
only by the strip FGDE. Since, within a turbine, expansion can be 
carried down to the lowest condenser pressures, turbines can utilize the 
large triangular expansion area, CHE, in addition to all that the engine 
gains. See the author's Steam Power Plant Auxiliaries And Acces- 
sories for further information. 

265. The Usual Vacuums Carried In Steam-turbine 
Practice Are as follows: (1) Where circulating water is not 



very yleniiful or where it must he pumped great distances: 27 
to 28 in. (2) Where circulating water is plentiful and always in 
large-capacity stations: 28 to 29 in. The smaller values are, in 
each case, the vacuums carried in the summer months; the 
lower values are those which are carried in the winter months. 
Although the values given above are quite commonly observed, 
the most economical vacuum should be determined for every 
plant before adopting a standard. This is done by a compari- 
son of operating costs with different vacuums. Higher 
average vacuums, and consequently more economical operation, 


15 16 .27 

Vacuum Referred To 30-In. Daromeier 


Fig. 250. — Showing typical relations of power-production cost to vacuum. This 
graph is plotted for specific conditions but the general principle which it illustrates is 
characteristic. For these conditions, the greatest decrease in net cost, at A, occurs 
with a vacuum of 28.6 in. mercury column. The net decrease in cost, BA, = (Decrease 
in fuel and feed-water cost, BD) — {Increase in fixed charges, maintenance, and circulating- 
water cost, BC). That is, to determine the locations of points along EK for each different 
vacuum, the corresponding vertical distance between EF and EH is subtracted from the 
vertical distance between EG and EH. 

are always possible in the northern than in the central and 
southern states. This is because of the lower temperatures 
of the cooling water in the northern states. 

Example. — In the plant and for the conditions for which Fig. 250 
was plotted, the most economical vacuum (at A) is 28.6 in. of mercury. 
With this vacuum the net cost of power production is 6.6 per cent, less 
than if only a 26-in. vacuum were carried. With more or less vacuum 
than 28.6 in., the net power cost would be greater, as shown by AK 
and EA. 

266. To Compute The Effect On A Turbine's Water Rate 
Of Changing The Vacuum, the factors given in Table 256 may 
be used whenever manufacturer's correction curves (Sec. 268) 


are not available. The values given in Table 256 are appli- 
cable only up to the vacuum at which the turbine was designed 
to be most efficient — generally 28.5 to 29 in. The method of 
applying these factors is illustrated in the following example. 

Example. — A certain turbine, when operating under a 27-in. vacuum, 
has a water rate of 12 lb. per kw.-hr. (The supply-steam pressure and 
the superheat — if any — may be any reasonable values.) What water 
rate may be expected from this turbine when operating under a 28.5-in. 
vacuum? Solution. — By Table 256 the water rate will be decreased 
6 per cent, by raising the vacuum to 28 in., and will be further decreased 
0.5 X 10 = 5 per cent, by raising the vacuum from 28 to 28.5 in. Hence, 
the -per cent, decrease = 6+5 = 11 per cent. Therefore, the water rate 
decrease = 0.11 X 12 = 1.31 lb., or, with a 28.5-in. vacuum, the actual 
water rate = 12 — 1.31 = 10.69 lb. per kw.-hr. 

6 8 10 II 14 16 16 lb 21 14 26 IB 30 32 34 36 56 40 
Back-Pressure On Turbine, Lb. Per 5q.ln.0age 

Fig. 251. — Graphs showing effects of increasing the back pressure on the water rates 
of non-condensing turbines. (-B. F. Sturtevant Co.) 

267. Increasing The Back Pressure On A Non-condensing 
Turbine Increases The Water Rate And Decreases The 
Thermal Efficiency (Fig. 251). — Since the back pressure on 
a non-condensing turbine corresponds exactly to the vacuum 
on a condensing turbine, all of the previous discussion con- 
cerning vacuums applies, to a greater or less degree, to back- 
pressures — the chief difference being in the magnitude of the 
effect of a given pressure change in the two cases. The graphs 


of Fig. 251 illustrate the effects on the water rates of increasing 
the back-pressure from atmospheric to different values and 
shows how these effects vary with different initial (supply) 
steam pressures. 

Note. — To Compute The Effect On A Non-condensing Turbine's 
Water Rate Of Changing The Back Pressure, the factors given in 
Table 256 or the graphs of Fig. 251 (which are more accurate) may be 
used. The method of using these graphs is illustrated in the following 

Example. — A non-condensing turbine which is operating with a supply 
pressure of 150 lb. per sq. in. gage (any reasonable superheat or no super- 
heat) and a back pressure of 10 lb. per sq. in. gage, shows a water rate, 
by test, of 44.8 lb. per hp.-hr. What water rate might the turbine be 
expected to have if the back pressure were raised to 25 lb. per sq. in. gage? 
Solution. — From Fig. 251, the water rate with 10-lb. back pressure and 
150 lb. per sq. in. supply pressure, is 25.5 per cent, higher than it would be 
with atmospheric exhaust. Hence, with atmospheric exhaust, the water 
rate = 44.8 ^ (1.00 + 0.255) - 35.7 lb. per hp.-hr. Also, from Fig. 251, 
the water rate with 25-lb. back pressure is 60 per cent, higher than with 
atmospheric exhaust. Hence, with 25-lb. back pressure, the water rate = 
35.7 + (0.60 X 35.7) = 35.7 + 21 A = 57.1 lb. per hp.-hr. 

268. Manufacturers Sometimes Supply Performance 
"Correction Graphs" With Turbines (Figs. 252, 253 and 254). 
The purpose of such graphs is to provide the purchaser with 
more accurate means, than the factors of Table 256, for com- 
puting the probable effects on the turbine's water rate of chang- 
ing the supply pressure, superheat, and vacuum. A very 
important application of such curves is for making " corrections " 
to the results of an acceptance test (Sec. 226) in which the 
exact steam conditions of the manufacturer's guarantee did 
not prevail. The use of performance correction graphs for 
verifying guarantees is explained in following Sec. 269. 

269. The Water Rates At The Steam Conditions Of An 
Acceptance Test May Be Corrected To The Water Rates 
Which Would Have Obtained If The Acceptance Test Had 
Been Made Under The Steam Conditions Of The Guarantee 
by the following formulas : 

(^^) <^ = (i-© + (i-w;) + (i-|;)('^--'^') 

(56) Wc = Wr - CWt (lb. per kw.-hr.) 


Wherein: C — the net correction factor, expressed as a decimal. 
Wg = the full-load water rate of the turbine at the steam 
conditions specified in the guarantee. Wf = the full-load 
water rate at the steam pressure of the acceptance test as 
determined from the pressure correction graph (Fig. 252). Ws 
= the full-load water rate at the temperature of the superheat 
of the acceptance test as determined from the superheat correc- 
tion graph (Fig. 253). Wy = the full-load water rate at the 
vacuum of the acceptance test as determined from the vacuum 
correction graph (Fig. 254). Wc = the corrected water rate; 
that is, the water rate after correction from the acceptance-test 
steam conditions to the steam conditions of the guarantee. 
Wt- = the water rate as determined by the acceptance test. 
All water rates are expressed either in pounds per kilowatt hour 
or per brake horsepower hour. 

Note. — The Net Correction Factor, C, For. (55), is the algebraic 
sum (see example below) of the individual correction factors that must 
be applied to correct for the change in the water rate which will be caused 
by a change in the steam pressure, superheat, or vacuum. In applying 
For. (56), it is assumed that the steam consumption at fractional loads 
will be changed by the same percentage as at full load for the same change 
in pressure, superheat, and vacuum. This assumption is, for all practical 
purposes, true within the range of from 50 to 125 per cent, of full-rated 
load. The method of application of these formulas is explained by the 
example below. 

Example. — A 500-kw., 3,600-r.p.m., turbo-generator was sold under 
the guarantee (Sec. 285) that when operating at rated speed at a steam 
pressure of 150 lb. per sq. in., gage, 50° F. superheat, and a 28-in. referred 
vacuum (Sec. 231), it will have the following water rates at the various 
loads : 

CONDITIONS = 150 lb. per sq.-in. gage; 50° F. superheat; 28-in. vacuum. 

Load in kw 






Per cent of rated load 


Guaranteed water rate 
kw -hr 

in lb. per 


When the acceptance test was made — at 50, 75, and 100 per cent, of 
the rated full load at the rated speed — under a steam pressure of 175 lb. 
per sq.-in. gage, 100° F. superheat, and a 27-in. referred vacuum, the 
turbine was found to have the following water rates : 



CONDITIONS = 175 lb. per sq.-in. gage; 100° F. superheat; 27-in. 


Load in kw 

Per cent, of rated load 

Water rate in lb. per kw.-hr. by- 
acceptance test 



The full-load correction graphs (Figs. 252, 253, and 254) for pressure, 
superheat, and vacuum corrections, are furnished by the turbine manu- 






















26-ln Vaccu 


E u 

50* F. Superheat. 

- 3 





















Full Loaol,500Kw.. 






























5 i 







r 2 







. . 






. _ 










. - 


























5tecim Pressure In Pounds Per Square Inch, Absolut e 

Fig. 252. — Graph for steam-pressure correction of a 500-kw. turbine. This graph 
shows the performance of the turbine at different supply-steam pressures but with the 
vacuum constant at 28 in. and with the superheat constant at 50°F. 

facturer for the particular turbine which is specified in the guarantee. 
What is the net correction factor? Correct the water rate for each of 
the various loads as determined under the steam conditions of the accept- 














7ft-ln Vnr 


Full Loaoi.500 Kw. 

^ 3 


E "rr 

vn ^ 









_ J 






















- - 












E o 

















Superheat In " F. 

Fig. 253. — Graph for superheat correction of a 500-kw. turbine. This graph shows 
the performance of the turbine at different superheats but with the supply-steam pres- 
sure constant at 150 lb. per sq. in. gage and the vacuum constant at 28 in. 

ance test to the conditions of the guarantee specification. 

Solution. — The specification guarantees a water rate of 17.4 lb. per 
kw.-hr. at fuU load, Wg of For. (55) = 17.4. From Fig. 252, Wp at 


175 lb. per sq. in., gage (189.7 lb. per sq. in., abs.) = 16.5 lb. per kw.-hr. 
From Fig. 253, W^ at 100°F. superheat = 16.7 lb. per kw.-hr. From Fig. 
254, Wv at 27-in. vacuum = 18.5 lb. per kw.-hr. Therefore, by substitu- 
tion in For. (55), the net correction factor, C, = (1 — Wc?/Wp) + (1 — 









- — 1 



DO r. ouperneaT. i 

t-- - 

150 Lb.fV.r Sn.In CnnpJ 








Full Load. 500 Kw. 

n 1 1 III 1 1 1 





: - 










E o 






E ^ 















Ycicuum In Inched Of Mercury 

Fig. 254. — Vacuum correction graph for a 500-kw. turbine. This graph shows the 
performance of the turbine at different vacuums but with the superheat constant at 
50° F. and the supply-steam pressure constant at 150 lb. per sq. in. gage. 

Wg/Ws) + (1 - Wg/Wv) = [1 - (17.4 - 16.5)]+ [1 - (17.4 ^ 16.7)] 
+ [1 - (17.4 ~ 18.5)] = (1 - 1.054) +'(1 - 1.042) + (1 - 0.940) = 
-0.054 - 0.042 + 0.060 = -0.036. 

The water rate, Wt, at full load as determined by the acceptance test 
is 16.5 lb. per kw.-hr. The net correction factor as determined above = 

2&-lnch Vacuum: 
50° F: 'Superheat 
150 Lb. Per 5q. In.: 
Gaqe Pressure 

200 300 400 

Output In Kilowatts 

Fig. 255. — Graphs showing: (1) The guaranteed water rate. (2) The water rate 
as corrected, from the steam conditions obtaining during the test, to the steam condi- 
tions on which the guaranteed water rate is based. 

—0.036. Therefore, by For. (56), the corrected water rate at full load, 
Wc, =Wt - CWt = 16.5 - (-0.036 X 16.5) = 16.5 + 0.6 = 17.1 lb. 
per kw.-hr. Similarly, it is found that the corrected water rate at 75 per 
cent, full load = 16.9 - (-0.036 X 16.9) = 17.5 lb. per kw.-hr. And, the 



corrected water rate at 50 per cent, full load = 18.9 — (—0.036 X 18.9) = 
19.6 lb. per kw.-hr. The following table shows the tabulation of the cor- 
rected water rates: 

Load in kw 

Per cent, of rated load 

Corrected water rate in lb. per kw.- 



By comparing the corrected water rates at the various loads with the 
guaranteed water rates at the corresponding loads, it is found that the 
water rates as determined by the acceptance test are lower than those 
which are guaranteed by the manufacturer. The water rates as deter- 
mined by test and those which are guaranteed by the manufacturer may 
be readily compared by plotting a graph of each, against the load in kilo- 
watts or brake horsepower. In Fig. 255, the graphs of the corrected 
water rates and the guaranteed water rates of this 500-kw. turbine are 
plotted against the loads in kilowatts. 

270. Water-rate Correction Graphs For Changed Pressure, 
Superheat Or Vacuum Applying To Any High -efficiency, Multi- 

-40 -20 ZO 

Change In Superheat -"F. 
Fig. 256 — 'Graph for superheat correction for turbine water rates. Supply steam 
pressure and vacuum are assumed to be constant. "This correction does not apply 
for superheats below 40° F." (Allis-Chalmers Mfg. Co., June 6, 1922.) 

stage, Impulse Or Reaction Turbine are given in Figs. 
256, 257 and 258. The accuracy of the results given by them 
will not be affected by the system of speed regulation which is 
employed on the turbine. These graphs are used in essentially 
the same manner as are those of Figs. 252, 253 and 254 except 
that these are more general in their application. 


An Explanation of the graphs of Figs. 256, 257 and 258, as quoted 
from a letter from the AUis-Chalmers Co. of June 8, 1922 is: "The 
graphs show the percentage change in steam consumption with changes 
in the steam conditions of an actual turbine installation. They do not 
apply if the turbine is altered in a way to render it more suitable for the 
changed conditions. These correction graphs apply only to the fixed 
ranges of steam conditions and loads which are, where necessary, speci- 
fied on them. This matter of limitations is important. It is not believed 
that it would be feasible to plot a set of usable correction graphs which 

+lb| 1 1 1 1 I 1 1 1 1 1 1 1 1 1 I 1 r - 


J. -Ti~- "" "" :: 

t ^ * Ay 

o -- " "^siw: '- 

«. .,^ ' i%3 it. " 

^ +^5 ^ M'^sL^'/, 

o ^s / ^'^i?/ ;; 

^ c ^ 4& I- rr^TA'; 

-1 "" ^^'^'u, FrNt? 

5 ,^ " '<yfn ■)■ ITS. 

B^+^o-- ykrB/J >s 

,Jo - -^-.^e *"> 

^Z ' ' '^^^ 

'* ^ik 

^c --------- --^^^^- 


O ' r " " 



■* » S 

!s^'iN-_- - - - 

^ ---------- - : 

_^>.;>^_ - - 

- >s^,. 

O C A 

:. __: !Sfc 

-^ i - 

:_^h, x- - 



" " " : - - --"*sS* - 

5i ::::::i::::::::::::::::± 

:::.:: s ^. 


1 1 II 1 IJsU 




^ -loMiiiinimiuiiiiMi Ml 

r^ e ^ J T tr\ I Q _A«-. 

Vacuum In Inches Mercury Column Referred To 30-ln. Barome+er 

Fig. 257. — Graph for vacuum correction for turbine water rates. Supply-steam, 
pressure and superheat are assumed to be constant. These corrections apply on most 
economical load and less only. (Allis-Chalmers Mfg. Co., June 6, 1922.) As defined 
by the AUis-Chalmers Mfg. Co.: A high vacuum turbine is one which, when tested with 
all conditions constant except vacuum, shows its best Rankine cycle ratio at a vacuum 
exceeding 27-in. mercury column referred to 30-in. mercury column. A low vacuum 
turbine is similarly one which has its best Rankine cycle ratio at a vacuum below 27 in. 
mercury column. 

would be reasonably accurate for all steam conditions. Such graphs 
would be too complicated. To insure simplicity the load limitation has 
not been applied to the vacuum correction graph (Fig. 257). However, 
it is a fact that the most economical load for a turbine will decrease as 
the vacuum is decreased and that the vacuum correction for an overload 
condition will be different from that for a normal load. In applying any 
correction graphs the most accurate method is to leave the test results 
without correction and to correct any guarantees or estimates of steam 
consumption to the steam pressure, superheat and vacuum which 
prevailed during the tests. The reason for this is that it has become 



apparent that the correction of test results to performance guarantee 
conditions has led to distortion in listing the performances of actual 

Example. — Showing the Application Of Pressure Correction 
Graph Fig. 258. This example is based on information furnished by 
E. H. Brown of the AlHs-Chalmers Company. A 5,000-kw. turbine unit 
which has its most economical load at 4,500 kw. was sold under a guaran- 
tee that when operating at rated speed at a steam pressure of 200 lb. 




u 5^000 





i2 40,000 





















.Tn-h/yl <,-f-enrn Jl-f- IKfi 1 >i 

Per 5cf. In. Gage- 





100° f: Superheat- 

Zd" Vacuum 










































m A-f?nO / h 1 



Per Scf. In. Gage 






100 F. Super hetT- 


. 2&"Vacc^un 















2,500 -5,000 ^.500 

Looiol, K i 1 w 01++5 




Fig. 257A. — Two graphs of the total steam consumptions, in pounds per hour, 
plotted against the load, in kilowatts, for the turbine unit given in the example of Sec. 
270. One graph is for steam at 200 lb. per sq. in., gage, and the other for steam at 150 
lb. per sq. in., gage. (From Allis-Chalrmrs Co., Graph No. St-1,398, September 7, 

per sq. in., gage, 100° F. superheat, and a 28-in. referred vacuum, it would 
have the following water rates at the various loads: (If the most econom- 
ical load for a given turbine is not known, it may be found by plotting 
the guaranteed water rates against the loads, to which they apply and 
then finding on this graph the lowest point. This point will correspond 
to the required — most economical — load.) 

Correct the water rate for each of the various loads given in the above 
table to a steam condition of 150 lb. per sq. in., gage, the superheat and 
vacuum to be the same as those specified in the guarantee. 


uJKTd+s j-o aniioA uq paiiddv ^g 01 uoipajJOD 4.033 jad 



Table 270A. — Guarantee Conditions 200 lb. per sq. in., gage; 100° F. 
superheat; 28-in. vacuum. 

Load in kw. 





Guaranteed water rate in lb. per 


Total steam — lb. per hr 





Solution. — Since in this problem the correction is to be made for a 
change in steam pressure, the correction factors must be taken from the 
pressure-correction diagram (Fig. 258). But the corrections for the 
given loads cannot be taken directly from this graph because the loads 
given in the guarantee are not the certain fractional loads (50, 75 and 100 
per cent, and greater) of the most economical load at the base steam pres- 
sure which are plotted on the graph. The corrections can be obtained 
indirectly from the pressure-correction diagram by the following method. 

(1) Plot the guaranteed total steam consumption per hour against the load 
in kilowatts. 

(2) From this graph find the guaranteed steam consumption per hour for 
the certain fractional loads (50, 75 and 100 per cent, and greater) of the most 
economical load at the base steam pressure, which are plotted on the pressure- 
correction diagram. 

(3) Find the guaranteed water rates at the base steam pressures for these 
fractional loads. 

(4) Read from Fig. 258 the corrections to be applied on the value of the 
water rate at the base steam pressure to obtain the value of the water rate at 
the test steam pressure. 

(5) By means of this correction calculate the water rate and total steam 
consumption per hour at the test steam pressure. 

(6) Plot the total steam consumption per hour at test steam pressure 
against the load in kilowatts. 

(7) From this latter graph read the total steam consumption per hour at 
test steam pressure for the given loads. 

(8) Calculate, from the total steam consumptions per hour obtained in (7), 
the water rates at the test steam pressure for the given loads. 

(9) Compare the economies of the turbine at the two steam pressures. 
The method outlined above, when applied to the solution of this 

example will result in the following procedure. 

(1) The guaranteed total steam consumptions per hour should be 
plotted against the given loads in kilowatts. The graph A, Fig. 257A, 
will result. Note that the graph consists of two straight lines with differ- 
ent slopes, the charge of slope occurring near the most economical load. 


(2) Since the most economical load at the base steam pressure is, in 
this example, 4,500 kw., the 50, 75, 100 and 111 per cent, loads are 2,250 
3,375, 4,500 and 5,000 kw. These will be used as the loads at which 
corrections will be made. The total steam consumptions per hour at 
guarantee conditions for these loads can then be read from graph A, 
Fig. 257 A, as tabulated in Table 2705. 

(3) The water rates at the base steam pressure for these loads must 
then be calculated by dividing the total steam consumption in pounds 
per hour by the loads in kilowatts. These values are given in line 3 
of Table 270B. 

(4) By following up vertically the —50 lb. change-in-steam-pressure 
line on the pressure-correction diagram (Fig. 258), the correction factors 
for the four loads may be obtained. Note that the first part of all of the 
correction curves for loads less than the most economical load coincide 
along the line marked ^^For loads less than the most economicaV^ and then 
they branch off from this line, the larger-load curves branching off first. 
The branching of the 50 per cent, load curve from this line is not shown 
on Fig. 258 as the diagram is not large enough. The values of these 
corrections as taken from Fig. 258 are given in Table 270B. 

(5) The water rates for these loads are then found by multiplying 
the base water rate by 1 plus the correction factor expressed as decimal. 
Thus for the load of 2,250 kw., the water rate at 150 lb. 'per sq. in., gage, 
steam pressure = 15.11 X [1 + (0.85 -^ 100)] = 15.11 X 1.0085 = 15.23 
lb. per kw.-hr., which checks with the value given in Table 2705. From 
the water rates thus obtained, the total steam consumption per hour at 
150 lb. per sq. in. gage pressure can be calculated by multiplying the 
water rate per kilowatt-hour at each load by the load in kilowatts. The 
values given in the last line of Table 2705 will result. 

(6) These steam consumptions per hour at 150 lb. per sq. in, gage, 
steam pressure should then be plotted against their respective loads in 
kilowatts. The graph B, Fig. 257 A, will result. 

(7) The steam consumptions per hour at test conditions for the given 
loads can then be read from the graph, by following up the vertical line 
corresponding to the load. The values of these consumptions are tabu- 
lated in the Table 270C, line 1. 

(8) By dividing the total steam consumption in pounds, per hour at 
150 lb. per sq. in., gage, steam pressure, for each given load by the 
load in kilowatts, the water rates in pounds per kilowatt-hour can be 
obtained. These are listed in line 2 Table 270C. 

(9) A comparison of the two water rates should be made to show the 
increase, in per cent., in the water rate. This can be done as follows: 
The water rate for a load of 2,500 kw. at 150 lb. per sq. in., gage, steam 
pressure, is 14.92 lb. per kw.-hr. (from Table 270C) and that for a steam 
pressure of 200 lb. per sq. in., gage, was guaranteed as 14.8 lb. per kw.-hr. 
The change from a pressure of 200 lb. per sq. in., gage, to one of 150 lb. 
per sq. in., gage, causes, an increase in the water, in per cent. = 100 



(14.92 - 14.8) -^ 14.8 = 12 h- 14.8 = 0.85 per cent. These values are 
listed in line 3 of Table 270C. 

270B. Table Showing Values Obtained During Correction To 
Conditions of 150 lb. per sq. in., gage; 100° F. superheat; 28-in. vacuum. 

Load in kw. 

2,250 3,375 4,500 5,000 

Load in per cent, of most eco- 
nomical at base steam pressure. 

Total steam, lb. per hr. at 200 lb. 
per sq. in., gage 

Lb. per kw.-hr. at 200 lb. per sq. 
in., gage 

Correction, in per cent., on 2001b. 
per sq. in. gage, steam pressure 
values (read from Fig. 258 at — 
501b. change in steam pressure) 

Lb. per kw.-hr. corrected to 1501b. 
per sq. in., gage, steam pressure. 

Total steam at 150 lb. per sq. in., 
gage, steam pressure 










+ .85 

+ 1.3 













270C. Table Showing Comparison of economy at 150 lb. per sq. in. 
gage, steam pressure, with that at 200 lb. per sq. in gage pressure. 

Load in kw. 

2,500 3,750 4,500 5,000 

Total steam at 150 lb. per sq. in., 

Lb. per kw.-hr. at 150 lb. per sq. 
in., gage 

Per cent, increase in steam con- 
sumptions in change of steam 
pressure from 200 lb. per sq. in., 
gage, to 150 lb. per sq. in., gage. 









The same method may be used where the steam pressure is increased, 
and if necessary, correctons for superheat, vacuum and steam pressure 
may all be applied to one value of economy. 

Note. — The ''Base Pressure" And "Superheat" Are Those From 
V/hich Values Are To Be Corrected. — If the guarantee water rates 
are to be corrected to test conditions — as is recommended in the preceding 



quotation — then the pressure and superheat values which are stated in 
the guarantee are the "base" values. If, however, the test results are to 
be corrected to guarantee conditions — as is done in the example under 
Sec. 269 — then the pressure and superheat values which obtained during 
the test become the "base" values. 


1. Upon what do the water rates and efficiency of a turbine depend? State the 
relation in general terms. 

2. State the formulas which give the theoretical water rate and thermal efficiency of 
any turbine. 

3. Does the thermal efficiency of a turbine increase in the same proportion as the 
water decreases when the supply conditions are varied? Give some values to prove this. 

4. State what factors determine the theoretical water rate and thermal efficiency 
of a turbine. What other factor affects the actual water rate and thermal efficiency? 

5. State the principal forms in which losses occur in steam turbines. What property 
of the steam largely affects the amount of the losses? 

6. Why are the percentage losses greater in small turbines than in large ones? 

7. Are turbines designed for specific steam conditions? How does this fact affect 
their operation? Explain fully. 

8. What would be the action of the steam in a turbine if it were operated under 
steam conditions much different from those for which it was designed? Explain fully. 

9. What effect is produced on the capacity of an existing turbine by increasing its 
supply steam pressure, superheat, or vacuum? 

10. State the approximate factors for calculating the change in water rate due to 
changes of supply pressure. Superheat. Exhaust pressure. 

11. What is the effect on a turbine's efficiency of increasing the supply pressure? 
Explain fully. 

12. What steam pressures are most advisable for turbine operation? 

13. How would you compute the effect on a turbine's water rate of changing the 
supply pressure? 

14. What is the effect on a turbine's efficiency of increasing the superheat of its supply 
steam? Explain why. 

15. What superheats are most advisable for turbine operation? 

16. What fuel saving may be expected from superheating? Why is very high super- 
heat not economical? 

17. How is the most economical superheat for a given plant determined? Draw a 
typical set of graphs to illustrate the principle. 

18. How would you compute the effect on a turbine's water rate of changing the 
superheat of the supply steam? 

19. What effect has the vacuum on the efficiency of a steam turbine? Are there any 
practical limits? 

20. Explain why turbines can more effectively utilize high vacuums than can steam 
engines. Draw the pressure-volume diagrams for the two classes of machines. 

21. What are the usual vacuums that are carried in turbine plants? 

22. How is the most economical vacuum for a given plant determined? Draw a 
typical set of graphs to illustrate the principle. 

23. How would you compute the effect on a turbine's water rate of changing the 

24. What is the effect on a non-condensing turbine's water rate and thermal efficiency 
of changing the back pressure in the exhaust pipe? How would you compute the effect? 

25. What are performance correction curves? For what are they used? 

26. Explain how you would correct the results of an acceptance test to the conditions 
of the guarantee? Explain fully. 


271. Steam-turbine Economics Is To Be Understood To 
Mean the study of the operating costs (see note below) of steam 
turbines. The purpose of such studies may be: (1) To deter- 
mine the cost of energy, so that it may be known at what price 
it may be profitably sold or that the management may know 
what the energy is costing. (2) To determine the most desirable 
turbine for a new plant or for addition to an existing plant. 
(3) To determine whether a turbine is more desirable than a 
prime mover of some other type. 

Note. — The Operating Costs Of Any Machine are generally 
grouped into two classes: (1) The Fixed Charges, Sec. 272, which are 
those expenses that are incidental to the oivning of the machine; the fixed 
charges include: (a) /n^eres^ on invested capital. (6) Depreciation, (c) 
Taxes and insurance, (d) Rental and office expense. (2) The Oper- 
ating Charges, Sec. 273, which are those expenses that arise when the 
machine is operated; they include: (a) Labor and attendance, (b) Fuel 
and water, (c) Repairs and maintenance, (d) Supplies, such as waste, 
oil, and the like. For a more thorough treatment of operating costs, 
see the author's Steam-engine Principles and Practice. 

272. The Annual Amount Of The Fixed Charges For 
Turbines varies from about 11 to 15 per cent, of the first cost 
of the turbine and auxiliaries (installed). The exact percent- 
age to be used in any given case can be determined by taking 
the sum of: (1) The current interest rate. (2) The depreciation 
rate, which is generally assumed as 5 per cent. (3) The 
tax rate. (4) The insurance rate. (5) The rental and office 
expenses which are chargeable to the turbine, expressed as 
percentage of the first cost of the turbine. For the purpose 
of good bookkeeping, the interest and rental should, rightfully, 
always be charged against the operation of the turbine whether 
it is actually paid out or not. In this way only, can the tur- 



bine be properly compared with other equipment which is 
more costly or which occupies a greater amount of space. 

Note. — The Fixed Charges Are So Called Because their amount 
is the same regardless of whether the machine is in operation or not. In 
this way they differ, as will be shown, from the operating charges which 
increase with the output of the machine. 

Example. — A turbine installation cost $20,000. If money can be 
borrowed at 6 per cent., the tax rate is 13^ per cent., the insurance rate 
is I'i per cent., and if the rental and office expenses amount to $400 per 
year, what is the annual amount of the fixed charges? Solution. — The 
amount of the rental and office expense is 400 -h 20,000 = 0.02 = 2 per 
cent. Assume that depreciation is 5 per cent. Hence, the annual fixed 
charges = $20,000 X (6 + 5 + 1.5 + 0.5 + 2) -^ 100 = 20,000 X 0.15 = 
$3,000. Hence it costs the owner of this turbine $3,000 a year merely 
to own it, whether or not it is operated. 

273. The Unit Operating Charges Of Turbines Vary Widely 
And Depend On Many Things ; see following note and 
Table 274. Reviewing the items (note under Sec. 271) 
which constitute the operating charges to note how these 
items may vary, it follows that: (1) The unit labor and 
attendance expense will vary with the size of the plant and 
the load which the plant carries because one attendant can 
generally care for the generating unit whether it has large 
or small capacity or whether it runs at full or partial load; 
also, very frequently one attendant can just as easily care 
for several machines as for only one. (2) The unit fuel and 
water expense depends upon the efficiency of the boiler, the 
cost of the coal and the method of handling and firing it, the 
water rate of the turbine, the quantity of cooling water 
required if any, the cost of water or the distance it must be 
pumped. (3) The unit maintenance and repair expense 
depends on the amount of repairs or maintenance which are 
necessary and upon the output of the machine. (4) The cost 
of supplies varies somewhat but, since this item is always 
small, it is unnecessary to dwell upon it at this point. 

Note. — The Unit Charges For Turbines Are found by dividing 
the total charges over a certain period of time by the number of energy 
units which are produced during that period. Unit charges are generally 
computed on a yearly or monthly basis and on the basis of kilowatt-hours 
or horsepower-hours produced. The sum of the several unit charges is 
called the unit operating cost. 

Sec. 274] 



274. Table Showing Operating Charges For Two Power 
Plants in a given month as taken from the records of the 
operating company; station A consisted of ten 500-hp. boilers 

Station A 

Station B 

Kw.-hr. generated 

Tons of coal 

Tons of ash 

Lb. water evaporated 

Lb. water evaporated per lb. coal. 

Lb. coal per kw.-hr 

Lb. water per kw.-hr 

Gal. engine oil per 1,000 kw.-hr. . . 
Gal. cylinder oil per 10,000 kw.-hr 



555 . 00 







322 . 10 







Total operating charges, in dollars, and operating charges per kw.-hr., 

in cents 






Repairs : 

Dynamos and appliances .... 



Pumps, pipes, fittings, and 


Operating boilers 

Operating engines and dynamos 



Lubricants and waste 

Miscellaneous expense 

Total, except fuel 


Coal labor, car to boiler room . . 

Total cost 

Average cost of coal on floor of 
boiler room 






693 . 66 



220 . 12 




198 . 62 






























with hand-fired furnaces, no coal-handUng apparatus, burning 
Illinois screenings, and having 5,000 hp. of reciprocating engines; 
station B was a modern steam-turbine plant with coal- and 
ash -handling apparatus, economizers, superheaters, and also 
burning Illinois screenings. (From Gebhardt's Steam Power 
Plant Engineering.) 

275. The Unit Operating Cost For A Turbine Depends On 
The "Load Factor" (Fig. 259). The load factor is the ratio 
of the average power delivered by the turbine over a certain 




50 60 70 SO 
Load Fac-tor- Per Cent 

Fig. 259. — Graphs showing how load factor influences the cost of generating energy. 
Costs at switchboard for a 7,500-kw. steam electric central station. This is from Geb- 
harts, Steam Power Plant Engineering. 

time period to the maximum power-demand imposed on the 
turbine during that tinie period. That is: 


^ , - . Average power 

Load factor = ^^ — . 1 — — ^ 

•^ Maximum demand 


I^oad factors are expressed as: (1) Daily load factors. (2) 
Weekly load factors. (3) Yearly load factors. In Fig. 259, the 
yearly load factor is used. As is shown by Fig. 259, the total 
yearly amount of the fixed charges is independent of the load 
factor whereas the total operating charges increase as the load 
factor increases, but not directly. Also, the unit fixed charges 


and unit operating charges decrease as the load factor is 
increased. Hence, the unit operating cost varies very widely 
with different load factors. For a more complete discussion 
of load factor, demand factor and similar quantities see the 
author's Central Stations. 

Example. — If a plant generates 2,400 kw.-lir. of energy during a 24-hr. 
period and the maximum demand during that period is 150 kw., what is 
the load factor for this period? Solution. — Average power = kw.-hr./hr. 
= 2,400/24 = 100 kw. Hence the load factor = Average power /Maxi- 
mum demand = 100/150 = 0.675 or 67.5 per cent. 

Note. — The Lower The Load Factor, The Greater Will Be The 
Required Capacity Of The Generating Equipment, For A Given 
Average Load. If the probable energy required of a plant during a 
given period is known and the probable load factor is also known, then 
the probable maximum demand which will be imposed on the generating 
equipment can be computed thus: 

Example. — A plant must generate 500,000 kw.-hr. each month. The 
probable monthly load factor is 60 per cent. What will be the maxi- 
mum demand on the plant? In other words what maximum power 
output must the generating equipment be capable of handhng? Solu- 
tion. — Maximum demand = Average power/Load factor = 500,000 -J- 
(24 X 30)/0.60 = 1,116 kw. 

276. The Operating Costs Of Turbines Are Generally 
Computed And Included Together With Those Of The 
Boilers. — This is done because it would be very difficult, if not 
impossible, to determine specifically the fuel expense which 
is properly chargeable to the turbine. Hence, no attempt is 
generally made to separately determine the costs of the 
turbine. Instead, the operating cost of the entire plant is 
generally computed by adding together the boiler-room and 
turbine-room operating costs. The unit operating cost is then 
determined for the entire plant. This unit operating cost is 
then useful for comparison between the turbine plant and a 
steam-engine plant or an internal-combustion-engine plant. 

277. In Selecting A Prime Mover For Any Given Service, 
consideration must be given to the following factors: (1) 
Adaptability; that is consideration must be given to the dis- 
tinctive advantages and disadvantages, see Table 287, of the 
various plants which are being investigated. (2) Reliability. 


(3) Economics; that is, the operating costs (Sec. 271) of the 
various plants must be studied. In the following sections, 
the above factors will be discussed principally as they apply 
to steam-turbine selection. Also, since the selection of a steam 
turbine generally involves a decision between a steam engine 
and a turbine, the following discussion wiU treat principally of 
the relative merits of these two prime movers. 

278. To Render The Steam Turbine Adaptable To Various 
Services has been the aim of turbine engineers during recent 
years. Formerly turbines were only designed to run at very 
high rotative speeds (several thousand revolutions per minute) 
and hence could be used only with reduction gears to drive 
relatively high-speed machinery such as electric generators. 
Today, however, turbines are designed for rotor speeds as 
low as 1,200 r.p.m. and, with reduction gears, are being used 
to drive even the slowest-speed machinery. Inherently, 
however, the turbine is best adapted for driving high-speed 
machinery which must operate at a constant rotative speed. 
Hence, its most extensive use is for driving electric generators, 
centrifugal pumps, blowers, and like high-speed machinery. 
Furthermore, as has been shown in Div. 9, the turbine is 
adapted for almost any steam pressures and can be operated 
condensing or to exhaust against back pressures. 

Note. — The Steam Turbine Is Not Reversible And Cannot Be 
Efficiently Operated At Variable Speeds. — These two limitations 
are practically the only ones which need ever rule out the turbine from 
the viewpoint of adaptability. However, even these have been somewhat 
overcome in marine practice where, for reversing, a separate turbine is 
employed and, to secure maximum efficiency, full speed is maintained 
whenever possible. 

279. Modem Turbines Are Very Reliable. — Because of the 
small number of bearing surfaces in a turbine and because of 
its purely rotational motion, the lubrication of the bearings 
can be made very positive, Div. 10, and the wear is inappreci- 
able. If kept in proper alignment and carefully operated, a 
steam turbine is more reUable than a prime mover of any other 

Sec. 280] 




f— ^ 







■^^ — ^ 

















1— 1 































• "^ 


































• l-H 

















1— 1 





































&f CO 

2 § 

5? S 


a > 

o ^ 2 
•c -^ - 

03 > 
U > 

U -^ 

o ? t- 
•2 -^ ^ 

(U '. •« 
cS J* X 

^ £ ^ 

.2 -« 

►I a.S 












lO • ■ • 










o • • • 












(N • • • 


O »0 T)< lO (N 


tH CO CO ■* "O 

J> O CO LO o o o »o 
r-T T-H i-T ci (N M oo 












































































































































































Oi0O»CO00O00C0 00Ot^Tt<»-i05 







O iC o »o 

,-1 ,-1 l-H CO CO CO CO 


281. The Efficiency Or Steam Economy Of A Turbine 
Depends Principally On Its Size And Steam Conditions. — 

The effect of size is partially illustrated in Figs. 264 and 265; 
turbines of larger capacity than those represented in these 
graphs show even better efficiencies — the very large condensing 
turbines have water rates of about 11 lb. of steam per kw.-hr.; 
see Table 280. The effects of steam conditions — pressure, 
vacuum, and superheat — have been discussed in Div. 13. 
There seems to be little difference, if any, between the effici- 
encies of impulse and reaction turbines of equal capacity; re- 
action turbines, however, are not practicable in sizes smaller 
than about 125 kw. 

282. The Efficiency Or Steam Economy Of Turbines At 
Fractional Loads (Fig. 260) is very much better than that of 
engines. Figure 260 shows that the steam rate increases 
more as the load is decreased with small turbines than with 

^ large ones. The high efficiency 

of turbines at light loads is par- 
ticularly advantageous in electric 
power stations where turbines 
must frequently be operated at 
fractional loads so as to be ready 
for a sudden increase in station 

S; " 25 50 75 lOO 125 

'^ Per Cent Of Rated Full Load 

Fig. 260. — Graphs showing approxi- NoTE. — ThE CAPACITY RATING Of 

mate variation of the steam consump- ^ TURBINE GENERALLY MeaNS VeRY 

tion of turbines with variations of LiTTLE.-Turbines often are most effi- 
cient at loads which are considerably 
less than their rated capacity and are usually capable of supplying 
considerably more power than their rating. Large turbines are often 
rated at the maximum load which their generators are capable of devel- 
oping continuously (see Fig. 273). But this, too, is not always the basis 
of the rating. Hence, the meaning of a turbine rating is often quite 




pmall lur bines J 







' inn 

^r^^ — 



Large Turbine 

(About 15.000 / 

' ^ 






283. The Economy Of Low- And Mixed-pressure Turbines 

is, generally speaking, better than that of turbines of any 
other class. This is because that, with these turbines, the 
capacity of an existing plant can be greatly increased without 
increasing the fuel consumption. By utilizing the exhaust 

Sec. 284] 



steam from non-condensing engines, pumps, or other equip- 
ment, the capacity of a plant can often be increased by 80 
to 100 per cent, without any increase of the boiler capacity. 
Where condensing engines are in use, these may be run non- 
condensing and their exhaust then utilized in a turbine — an 
increase in capacity of 40 to 50 per cent, may thus be obtained 
with but a slightly greater amount of steam consumed. See 
Div. 9 on low- and mixed-pressure turbines. 

Note. — The Use Of Separate High-pressure Non-condensing And 
Low-pressure Turbines Is Not Advisable; the very-large capacity 

ITo+a I 





To-tal Energy In A Given Quantity Of 5+eatn Available 
For Heating And Power 


Available For Heating 



Bearing Fricfion And 


■ Converted Info Power | 

^Consumed by Auxiliaries ^^^ = Lost To Condenser 

- Energy Available for Heating 

Fig. 261. — Chart showing approximately the disposition of the heat energy in a given 
quantity of steam when it is used in turbines of different types. The bleeder turbine 
operation {III), can, on a moment's notice, be changed to either that of the non-con- 
densing {IV) turbine or any condition intermediate between II and IV — as power and 
heat requirements may demand. 

compound units, Sec. 68, are considered as being single units. Such an 
arrangement, although efficient in its use of steam, is not commercially 
economical because it necessitates a duplication of turbine and generator 
units — it is usually found that one high- pressure condensing turbine is 
better. Exhaust-steam turbines should, therefore, only be employed 
where profitable use can be made of the exhaust steam from existing 
steam-using equipment. 

284. The Economy Of Bleeder Turbines, Fig. 261, (see 
also Div. 9) lies in the fact that, by them, low-pressure steam is 
made available for heating or industrial services after the 
steam has been first used very efficiently to generate electrical 
energy in the bleeder turbine unit. By so arranging the load 


on a bleeder turbine that the turbine always consumes (re- 
ceives) considerably more steam than is extracted from it 
the turbine can thereby be made more efficient in its use of 
steam than would be a non-condensing turbine which consumed 
only the amount of steam that is necessary for low-pressure 
heating or the like. Bleeder turbines are, therefore, being 
used more and more as house turbines in large power stations — 
the auxiliaries being driven largely by electric motors which are 
supplied with energy from the generator which the bleeder 
turbine drives. 

Note, — For The Most Economical Application Of A Bleeder 
Turbine in an electrical generating station, it should be operated in 
conjunction with another (condensing) turbine. The total load is 
divided between the two units. The load on the bleeder turbine can 
then be changed, from time to time, as is necessary to insure that this 
turbine will always ''bleed" sufficient low-pressure steam to satisfy 
feed-water or other heating requirements. 

285. To Predict The Steam Rate Of A Contemplated 
Turbine, the method of Sec. 15 may be used for the first esti- 
mate ; or it may be read from Table 280. The exact water rate, 
however, can best be determined by applying to various 
manufacturers for their guarantees. Manufacturers generally 
specify steam economies which their turbines will actually 
exceed by a slight amount. This they do to be on the safe 
side. Having the builders' guarantees one may then make 
his final calculations. When bleeder or mixed-pressure tur- 
bines are contemplated, their low-pressure steam rates must 
very often be estimated; hence, undue accuracy in their water- 
rate calculations should be avoided. 

286. The Relative Economies Of Steam Turbines And 
Steam Engines depend, to a great extent, upon local conditions. 
Because they generally operate under different conditions it 
is often difficult to make reasonable comparisons between the 
two. Certain items of economy, however, are quite general 
in that they hold for nearly all comparisons — these items have 
been included in Table 287. Since the ffi'st cost of turbines is 
less than that of engines of equal capacity, the interest, taxes, 
insurance and depreciation charges are correspondingly less. 
The rental charges are also less, because of the fact that the 

Sec. 287] 



turbine occupies less space; see Figs. 262 and 263. Likewise 
with the other economy items given in Table 287. Practically 
the only item of economy which is not given in Table 287 is 

Horizontal Corliss' 

Fig. 262. — 'Comparative floor 
space occupied by steam engines 
and turbines. 

. Fig. 263. — ^Comparative head 
room necessary for steam engines 
and turbines. 

that of operating efficiency or steam economy; this item is 
treated in Sees. 288 and 289. 

287. Table Of Advantages And Disadvantages Of Steam 
Turbines And Steam Engines. 





Low first cost. 

Greater first cost. 

Low maintenance and attendance. 

Greater maintenance and attendance. 

Economy of space and foundation. 

Requires more space and larger foundation. 

Clean exhaust steam. 

Oil in exhaust steam. 

No vibration due to reciprocating parts. 

Reciprocating parts cause vibration. 

Uniform angular velocity. 

Angular velocity varies during each stroke. 
Heavy flywheel required. 

High efficiencies for, large variations in load. 

Decreased efficiency at fractional loads. 

Can utilize steam at high temperatures. 'High temperatures give trouble. 



Cannot be made reversible. 

Can be made reversible. 

Speed too high for many services. 

Runs at low angular speed. 

Runs at constant speed. 

Can be run at variable speeds. 

Condenser requires much water. 

Condenser requires less water. 


288. The Relative Steam Economies Of Non-condensing 
Turbines And Engines are illustrated in Fig. 264 for full-load 
operation; see also Table 280. It is well to note that the non- 
condensing turbine is not as efficient as the non-condensing 
engine. However, at fractional loads (Sec. 282), the turbine's 
efficiency is more nearly equal to the engine's. As is shown by 
Fig. 264, the efficiency of the turbine in the larger sizes is also 
more nearly equal to that of the engine than in the smaller 











Initial Wre55ure = lbO Lb. Her Sc^-ln. Gage 
Back Pressure = 1.5 Lb. Per 5a. In. Goae 






' 1 1 






t \ 













1 1 irii 1 5+pi-i 

m T 1- 



igh Speed, 

oinqie-Yalve Fnninc 




1 r 1 r-- — r- 


-_ _ 

_ __ 



orlbs F6ur-Voilve_ Engine - 








J 20 


."5 14 

^ . 

















Capoci+y In drake Horsepower 
Fig. 264. — Graphs showing comparative steam economies at full load of average non- 
condensing steam engines and steam turbines. These graphs show the water rates, in 
pounds per brake horsepower- hour, for engines and turbines supplied with dry satu- 
rated steam at 150 lb. per sq. in. gage and exhausting against a back pressure of 1.5 lb. 
per sq. in. gage. 

sizes. Furthermore, the turbine is more efficient with high 
steam pressures whereas the steam engine is more efficient 
with lower steam pressures. Although, as shown above, the 
efficiency of the non-condensing engine exceeds that of the 
non-condensing turbine, this is not to be taken to mean that 
the overall econornies are so related. Because of the turbine's 
lesser first cost, attendance, and maintenance expense, and 
because of its other advantages (Table 287), the turbine is, 
in many cases, more economical than the more efficient steam 

Sec. 289] 



288A. Turbine Steam Rates Are Also Less Likely To 
Increase With Years Of Service Than Are Engine Steam 

Rates. — This is because the only wearing parts of the turbine 
are the bearings, nozzles, and blading. The nozzles and blad- 
ing do not ''fit tight," even when the turbine is new. A small 
amount of wear, due to steam erosion, of these nozzles and 
blades will not produce excessive steam leakage as will a small 
amount of wear on engine valves or cylinders. 

3 lA 


Initial Pre55ure = 190Lb.Pe 

>r 5q. In. Gaqe 

Superheat^ ( :\15°F. 

Vriruiim rnnlnp*^ = 7ft-ln 



Vacuum, Turbines = 28^-ln. 



- -3600-R.P.M. Condensing Steam Turbines 





.-Compound Con 


Corliss Engines 









































Fig. 265. — Graphs showing comparative steam economies at full load of condensing 
engines and turbines. These graphs are based on steam supplied at 190 lb. per sq. in. 
gage and 125° F. superheat and on a vacuum for the engines of 26 in. and for the turbines 
28.5 in. The condensing steam turbine is more efficient than the compound condensing 
steam engine in capacities of J, 000 hp. and larger. 

289. The Relative Steam Economies Of Condensing 
Turbines And Engines, at full load, are illustrated in Fig. 265 ; 
see also Table 280. The efficiencies of the engines are seen to 
be better in the capacities below 1,000 hp., whereas in larger 
capacities the turbines show the better efficiencies. This 
comparison, it should be noted, is made with a greater vacuum 
on the turbines than on the engines; this is done because the 
turbine is most economical at greater vacuums than is the 
steam engine; Sec. 264. The turbine, therefore, requires more 
cooling water than does the engine and is less desirable where 
only a limited supply of cooling water is available or where 
the circulating water must be recooled in ponds or towers. 


290. Table Showing Applicability Of Steam Turbines And 
Engines In Units Of Small Capacity. — This table is based on a 
paper by J. S. Barstow before the A.S.M.E. in Dec, 1915 and 
applies chiefly to units of 500-hp. or less capacity. 



Condensing Units, Direct-connected 

(a) 60-cj/cZe generators in all sizes, 
(fe) 25-cycle generators above 1,000-kw. 

(c) Centrifugal pumping machinery oper- 

ating under substantially constant 
head and quantity conditions and 
at moderately high head, say from 
100 ft. up, depending on the size of 
the unit. 

(d) Fans and blowers for delivering air 

at pressures from 13^^ -in. water col- 
umn to 30 lb. per sq. in. 

Non-condensing Units, Direct-con- 
nected For All The Above, Pur- 
poses, In Those Cases Where: 

(a) Steam economy is not the prime fac- 

tor or the exhaust steam can be 
completely utilized. 

(b) Oil-free exhaust steam is desirable or 


Geared Units, Either Condensing 
OR Non-condensing, for all the above 
applications and, in addition, many 
others which would otherwise fall in 
the category of the steam engine, on 
account of the relatively slow speed of 
the apparatus to be driven. 

1. Non-condensing Units, Direct-con- 

nected Or Belted For: 

(a) Electric generators of all classes, ex- 

cepting exciter sets of small capac- 
ity unless belted from the main 

(b) Centrifugal pumping machinery op- 

erating under variable head and 
quantity conditions and at rela- 
tively low heads, say up to 100 ft., 
depending on the capacity of the 

(c) Pumps and compressors for deliver- 

ing water or gases in relatively 
small quantities and at relatively 
high pressures — in the case of 
pumps at pressures above 100 lb. 
per sq. in., compressors above 1 lb. 
per sq. in. 

(d) Fans and blowers (inchiding induced- 

draft fans) for handling air in vari- 
able quantities and at relatively 
low pressures, say not over 5-in. 
water column. 

(e) Line shafts of mills, where the driven 

apparatus is closely grouped and 
the load factor is good. 
(/) All apparatus requiring reversal in 
direction of rotation, as in hoisting 
engines, and the like. 

2. Condensing Units, Direct-connected 

OR belted, for all the above 
purposes, particularly where: 
(o) The condensing water supply is 

(6) The water must be recooled and re- 

Sec. 291] 



291. The First Costs Of Steam Turbines of different capaci- 
ties are given approximately in Table 280 and Fig. 266. 
The values given here must be understood to be only indica- 
tions and subject to the influence of local Conditions and 
market fluctuations as it is impossible to give prices which will 
be even nearly correct for any length of time due to the rapid 
change of prices. These prices are not intended to be accurate 
at any future date but, they may, however, be used for pre- 
liminary estimates of power-plant cost as they show how the 
price varies with the size of the unit. This relationship 
remains practically the same regardless of the change in price. 
If at any future date the percentage change of the average 

— ' 









o '^ 


• — "^^ 



J »i^ 


- t-^T 










: I" 




«. , 



= ! SS«^ 







= "* 


, 1 





Rated Capacit y - Kllow at 1 5 

Fig. 266. — Showing approximate prices of turbo-generator units of different capaci- 
ties. (The dots represent prices quoted on condensing units. The crosses represent 
prices quoted on non-condensing units. Prices are as of spring, 1922. Some of the 
units are not equipped with direct-connected exciters, but in most cases the price includes 
exciter. The omission or addition of the exciter makes little difference in the price per 
kw. Condensers are not included.) 

price based on that given here (spring 1922) is known, the price 
of any unit at that date can be approximately found by multi- 
plying the price given here by that percentage and making this 
correction to the price here given. Whenever reasonably 
accurate prices are required they should be obtained from the 
manufacturers. The graph of Fig. 266 shows remarkably 
well how the cost per kilowatt decreases with increased size of 
the unit. It will be noted from Fig. 266 that the price per 
kilowatt decreases very rapidly with an increase in the size 
of the unit for units between 50 and about 1,000 kw. capacity. 
Above 1,000 kw. capacity, the price per kilowatt does not 



decrease very much for an increase in capacity. Above 
30,000 kw. capacity the price per kilowatt is practically 

292. The Steps To Be Taken In Selecting A Prime Mover 
For A Given Service are: (1) Determine the load factor, Sec. 
275, and the hourly load variation if possible; if it is not possible 
to accurately determine the load variation, then try to obtain 
the probable load variation from some similar plant. (2) 
Determine the 7naximum load; in new plants, the maximum 
load must often be estimated. (3) Select the most desirable 
capacities of units; this should be done with a view toward 
always operating each unit at its most economical load. 
Generally speaking, the fewer units in a plant the better, 
provided always that there is sufficient generating capacity to 
carry the maximum peak with the largest unit out of service. 
(4) Get costs and performance guarantees (Sec. 294) for the 
different units of each type which is being considered; this 
usually requires the making of tentative building and machin- 
ery layout drawings of the arrangements which are under 
consideration. (5) Calculate the unit operating costs for each 
type over a yearly period ; to do this, estimates of the operating 
charges must be made. (6) Tabulate the estimates and decide 
on the type of equipment which shows the smallest unit operating 
cost, or is otherwise most desirable. 

The method of selecting a prime mover is explained by the 
following illustrative example, which is taken from the 
National Electric Light Association Prime Movers Commit- 
tee's Report for 1921. 

Note. — The Values In The Following Example, As It Is Here 
Used, Are Intended To Illustrate A Method Of Procedure rather 
than to "prove in" or "prove out" any certain type or class of power- 
generating equipment. Obviously, the values of the different elements 
which comprise the total cost will vary in different localities. The costs 
shown are for the vicinity of New York City in the year 1921. It is 
only by thus preparing an accurate tabular comparison of the costs of 
energy, as developed by different types of equipment and under different 
conditions, that the most economical equipment and steam conditions 
for a given location can be determined. In the N.E.L.A. report, above 
referred to, an energy-cost comparative analysis is also given for 200-kw. 
plants which operate at load factors of 25 and 75 per cent. 

Sec. 292] 



Example. — It is desired to select the most economical equipment for 
a generating station which is to furnish electrical energy at the average 
power-output rates stated in Table I below. The following equipment 
is to be considered: (a) Uniflow engines, (6) high-speed counterfiow 
engines, (c) turbines, (d) Corliss engines, (e) Diesel oil engines and (/) 
semi-Diesel oil engines. The most adaptable steam pressures may be 
assumed as 175 lb. per sq. in. for all units except the Corliss engines for 
which 150 lb. is to be used. Costs are to be determined for non-condens- 
ing (atmospheric exhaust) and for condensing operation both with satu- 
rated steam and with steam of 100° F. superheat. The condensing 
engines are to operate with 26-in. vacuum; the turbines with 28-in. 

The cost of coal is to be taken at $7.00 per ton, delivered. The heating 
value of the coal is 13,500 B.t.u. per lb. The oil engines are to be supplied 
with an oil of 18,500 B.t.u. per lb. heating value which will cost about 
$3.00 per bbl., delivered. The maximum peak load, assumed to occur 
only occasionally, is 200 kw. The average 24-hr, daily demand is 
assumed to vary as follows : 

Table I. — The Loads and Their Duration 

1. Load, in kw. 

2. Duration of 
load, hours 

3. Kw.-hr. of 
energy generated 


Peak load 






















The plant is assumed to be located in a small town near an adequate 
supply of water of the proper quality for condenser or oil-engine cooling. 
Also, a railroad siding is adjacent for the delivery of coal or oil. 

Solution. — Proceeding as suggested in Sec. 292, the steps are as follows : 

1. Determine The Load Factor And Load Variation. — The load 

variation has been determined by comparison with similar plants and 

found to be as shown in Table I above. The load factor is found thus: 

Total kilowatt-hours generated _ 2,400 

Hours duration 

The average load = 


100 kw. 

Now, the load factor for this average load will, from For. (57), be: 
Average load _ 100 
M aximum demand ~ 200 

Load factor — 

0.50 or 50 per cent. 


2. Determine The Maximum Load. — The maximum load is given in 
the problem as 200 kw. 

3. Select The Most Desirable Capacities Of Units. — To provide 

Turbine Unit %. 
Foundation ^i_ 


Fig. 267. — Sectional elevation of the 300-kw. and also of the 400-kw. (total capacity) 
steam turbine generating stations the plan views of which are shown in following illus- 
trations. (N.E.L.A., 1921, Prime Movers Report.) 

1 T^"""- ^'^.'.'Circulatinq-wafer pipes 

u u 

Fig. 268.— Plan view of the 400-kw., total capacity (2-200 kw. units) steam turbine 
generating station. See preceding illustration for section. 

sufficient generating capacity with the largest unit out of service and 
yet to have only a small number of units, it is thought advisable to con- 

Sec. 292] 



sider and make calculations for (a) two 200-kw. and (6) three 100-kw 
generating units of each type. 

4. Get Costs And Performance Guarantees For The Different 
Units.— The building in all cases is assumed to be of brick construction. 


■Cjrculafing-water pipes 


269.-Plan view of the 300-kw., total capacity (3-100 kw. units), steam turbine 
generating station. See preceding illustration, Fig. 267, for section. 

Load In Kilowatts 

Fig. 270.— Average steam consumptions per kilowatt-hour for 200-kw. condensing 

steam units. 

In all cases except for the belted Corliss engines, the roof trusses are of 
steel. The station to house the belted Corliss engines is designed with 
wooden roof trusses and a central line of posts on account of the long 
span required. 

The layouts of the buildings and principal equipment for the turbine 
plants are given in Figs. 267, 268, and 269. In the N.E.L.A. report 


from which this example is taken, layouts are shown for all of the differ- 
ent plants which are considered. From such layouts, contractors can 
make estimates. The investment costs are tabulated in Table IV below. 
The steam equipment is found to require boiler capacities as follows: (a) 
For the non-condensing turbines and for the Corliss engines (both con- 
densing and non-condensing), two 200-hp. boilers. (6) For all other 
cases, two 150-hp. boilers. Proposals and performance specifications 
were obtained from 65 manufacturers and averaged by classes. The 
average steam (or oil) consumptions were plotted into curves of which 
Fig. 270 is typical. 

5. Calculate The Unit Operating Cost For Each Type. — To do 
this, the yearly operating costs are first found and later, from these, the 
unit operating costs are found. The annual fixed charge is assumed to 
be 15 per cent, of the total investment cost for all plants, this figure 
including interest, taxes, depreciation and both liability and fire insur- 
ance. The fuel costs are thus determined: (a) For the oil-engine -plants. 
The oil consumed per 24-hr. day was computed by multiplying each item 
of column 3, Table I, by the fuel rate at the load shown in column 1. 
(The fuel rate is read from the guarantee curve.) From the daily oil 
consumption, the animal fuel-oil cost can readily be obtained. (6) For 
the steam plants. The steam consumed per 24-hr. day was computed by 
multiplying each item in column 3 of Table I by the steam rate at the 
load shown in column 1. Thus, for the 200-kw. condensing turbine plant: 

Table II. — Steam Consumption, 200-kw. Condensing Turbine Plant 

Load, in kfv^., 
from Table I 

Water rate, 
lb. per kw.-hr.. 
from Fig. 270 



from Table I 

Steam con- 
sumed, lb. 




























Now, allowing 9,600 lb. per day for losses due to pipe radiation, drips, 
and like, 3,200 lb. per daj^ for the boiler-feed pumps, and 9,900 lb. per 
day for the condenser air and circulating pumps, the total daily steam 
consumption = 63,350 + 9,600 + 3,200 + 9,900 = 86,050 lb. 

Since, in this example, the load factor is 50 per cent., the boiler effi- 
ciency will be about 64 per cent. Also, from steam tables, the total 

Sec. 292 



heat of dry saturated steam at 175 lb. per sq. in. gage is 1,198 B.t.u. per 
lb. If a feed-water temperature of 200° F. is assumed, the heat of the 
liquid (from steam table) is 168 B.t.u. per lb. Hence, the B.t.u. absorbed 
per pound of steam = 1,198 — 168 = 1,030 B.t.u. Therefore, with coal 
of 13,500 B.t.u. per lb. heating value, and a boiler efficiency of 64 per 
cent., the evaporation = 0.64 X 13,500 ^ 1,030 = 8.39 lb. steam per lb. 
of coal. Therefore, the daily coal co7isumption = 86,050 -^ (8.39 X 
2,000) = 5.13 tons per day. Hence, at $7.00 per ton, the annual coal 
cost = 5.13 X 365 X $7.00 = $3,100. (See Table IV.) 

The annual labor cost is computed by assuming the required atten- 
dants and their probable salaries, thus : 

Table III. — Attendants Required and Salaries 

Class of employee 

Number required 



Salary, each, 
per month 

Chief engineer . . 
Watch engineers 




Thus, for the 200-kw. condensing plant the annual cost of labor and 
superintendence = 12 X [175 + (2 X 125) + (3 X 110)] = $9,060. Now, 
since the regular power-plant force of attendants can, ordinarily, attend 
to the making of repairs about the plant, $1,000 of the annual salaries 
may be charged to repairs leaving the annual charge for labor and super- 
intendence = $9,060 - $1,000 = $8,060; see Table IV. 

The annual costs of lubricants, miscellaneous supplies, and pumping 
cooling water (for the oil engines) were estimated; the estimated values 
are given in Table IV. The annual costs of repairs were figured at 4 
per cent, of the investment costs of the engine plants and at 3 per cent. 
in the case of turbine plants. Thus, for the 200-kw. condensing-turbine 
plants the annual repair cost was figured as 3 per cent, of the investment 
cost for the generating units, condensing equipment, boilers, and feed 
pumps. The investment cost for this equipment = $19,460 + 7,000 + 
13,000 + 1,500 = $40,960. Therefore, the annual repair cost = 0.03 X 
$40,960 = $1,229; see Table IV. 

Thus, the total annual cost of operation is the sum of annual fixed 
charges, fuel, labor and superintendence, lubricant, miscellaneous sup- 
plies, and repair costs. This sum gives a value of $38,228 as shown in 
Table IV. 

Therefore, the unit operating costs may be computed by the formula: 

Annual operating cost 


Unit operating or energy cost 

Energy units delivered per year 


Table IV.— Showino Investment and Operatino 

Load Factor 50 Per Cent. Steam Pressure 175 Lb. per Sq. In. Gage for All 

Engines. Cost of 13,500 B.T.U. Coal $7.00 Per Net 

Non-condensing steam prime movers 
Back-pressure atmospheric 







Real estate 

Brick building 

Generating units, delivered and erected. . 
Switchboards and street lighting trans- 

Electric wiring and ducts 

Piping complete 

Condensing equipment 

ioundations, exclusive of building 

Oil filters and tanks 

Railroad siding. ■ 

Boilers, delivered and bricked in 

Feed w'ater heater 

Feed pumps 

Steel stack and flues 

Motor-driven pump for cooling water 

10,000 gal. fuel oil storage tank 

Air compressor and tanks 

Intake for circulating water 

Total investment. 

Cost op Operation: 

Fixed charges 15 per cent, on investment 


Labor and superintendent 


Miscellaneous supplies 


Cost pumping cooling water 

Total operating cost 

Cost per kw.-hr. operation (cents). . . . 
Cost per kw.-hr. fixed charges (units) . 

Cost per kw.-hr. total (cents) 

Cost per kw.-peak (dollars) 

A — Steam Equipment Desioned 












S 2,500 











$ 2,500 



















Real estate 

Brick building 

Generating units, delivered and erected. . 
Switchboards and street lighting trans- 

Electric wiring and ducts 

Piping complete 

Condensing equipment 

Foundations, exclusive of building. 

Oil filters and tanks.' 

Railroad siding. 

Boilers, delivered and bricked in . 
Feed water heater 

Feed pumps 

Steel stack and flues 

Motor-driven pump for cooling water-. 

10,000 gal. fuel oil storage tank 

Air compressor and tanks 

Intake for circulating water 

Total investment 

Cost of Operation: 

Fixed charges 15 per cent, on i 


Labor and superintendent. . . , 


Miscellaneous supplies 

Cost pumping cooling water. . 

Total operation cost. 

Cost per kw.-hr. operation (cents). . . . 
Cost per kw.-hr. fixed charges (cents). 

Cost per kw.-hr. total (cents) 

Cost per kw.-peak (dollars). > 

B— St^am Equipment Designed 


t 2,500 



t 17,273 









$ 17,513 

2,500 S 2,500 
25,504 26,634 



t 42,200 

$ 16,670 









i 14,722 

$ 2,500 





S 2,500 




S 18,232 






f 46,108 





Sec 293] 



CoBT or A 200 Kw. Centbal Station 

Units Except Corliss Engines. Steam Pressure ISO Lb. per Sq. In. Gage for Corliss 

Ton Delivered. Cost of 18,500 B.T.U. OU »3.00 Per Bbl. Delivered 

Condensing steam prime movers 


Engines condensing to 26-in. vacuum, turbines condensing to 28-in. vacuum 1 

Uniflow engines 



Corliss engines 



























For Satorated Steam 



















































































4., 700 



























































































































































S 41,476 

S 42,793 

$ 44,102 

$ 44,925 

$ 38,228 

$ 40,658 

$ 47,068 

$ 47,991 

$ 38,407 

$ 38,060 

$ 41,186 

$ 38,971 




















2 33 





























For Stteau 

Sdperheated 100° F. 

% 2,500 

S 2,500 

$ 2,500 

$ 2,500 

$ 2,500 !$ 2,500 

$ 2,500 

$ 2,500 

















































































































( 19,268 

t 20,183 

$ 18,300 

$ 19,340 

$ 15,729 

$ 17,445 

$ 19,662 

$ 20,902 





























- 300 












% 41,400 

S 42,807 [l 43.674 

$ 44,999 

$ 38,731 

$ 41,511 

$ 47,825 

$ 48,598 



































Thus, for the 200-kw. condensing-turbine plant, the unit operating 
cost = $38,228 H- (365 X 2,400) = $0.0437 or 4.37 ct. per kw.-hr.; see 
Table IV. 

For comparison, the investment cost per kilowatt of peak load was 
also computed for each type of unit. Thus, for the 200-kw. condensing 
turbine plant, the cost per kilowatt of peak load = (total investment cost) -^ 
(kilowatts peak-load capacity) = $102,260 ^ 200 = $511.30. 

6. Tabulate The Estimates And Decide On The Type Of Equip- 
MENT.^The estimates are here tabulated in Table IV. 

From the preceding tabulation (Table IV) it is evident that the plant 
with three 100-kw. semi-Diesel engines shows the least unit energy cost 
(4.34 ct. per kw.-hr.) which is but slightly less than that of the plant with 
two 200-kw. condensing turbines when supplied with saturated steam 
(4.37 ct. per kw.-hr.). Because of the lesser investment cost of the tur- 
bine plant and because of its greater reliabiUty, it would probably, for 
the stated conditions, be chosen in preference to the oil-engine plant. 

" Contrary to what seems to be the general belief, the lower steam rate 
which obtains with superheated steam is, in practically all cases, offset by 
higher fixed charges and fuel costs; and, except in the case of turbines, no 
net gain is realized by operating the plants condensing." 

293. The Information Which Should Be Given The Turbine 
Manufacturer When Requesting A Quotation is as follows: 
(1) What is wanted; turbine, turbo-generator, turbine-driven 
centrifugal pumps, etc. (2) Capacity; horsepower, kilovolt- 
amperes, kilowatts, or gallons per minute; always, if possible, 
for an alternating-current generator, state the power factor. 
(3) Speed; this need not generally be given if the driven 
machine is to be included in the quotation. (4) Steam condi- 
tions; boiler pressure, superheat, and back-pressure or vacuum. 
If a mixed-pressure or bleeder turbine is wanted give also the 
quantity and pressure of the low-pressure steam which is 
available or to be extracted. (5) Output conditions; whether 
alternating-current or direct-current generator is wanted, 
voltage, number of phases and frequency or head against which 
pump must discharge, etc. If an a.-c. generator is required 
state whether the exciter is wanted direct-connected on the 
main-turbine shaft or whether separate turbine-driven exciter 
is wanted. (6) Nature of load on driven-machine or on 
turbine; state whether load is composed largely of motors or 
whether it is principally a lighting load and also whether the 
load is steady or variable. 

Sec. 293] 





Milwaukee, Wisconsin, U. S. A '^.^.A. 


■ Jdduss St . louls. Mo. 

Allis-Chalnurs Manufacturing Company, hfrfinafur calUd the Company, proposes to furnish 
tht Purchaser, on the lollowing conditions, the machinery described below, or in the Company*! 
specifications attacked, which are made a part of this proposal, t. o. b. cars point of shipment. 

One (l_)___750-k;w. , at 80 per cent^ maxitmira rated 

turbo-alternator urtltoomi^ expansion Joint bnt not 

Including _exolter, as per attaohed specifications pages 5 to 9 
InoluslTo. _ .. 

All machinery (hall be insuUed by and at the expense of the Purchaser, unless otherwise expressly 
stipulated herein. 

The Company will repair f. o. b. works where made, or furnish without charge f. o. b, its works, a 
similar part to replace any material of its own manufacture which, within one year after shipment, is 
proven to have been defective at the time it was shipped, provided the Purchaser gives the Company Im- 
mediate written notice of such alleged delects. The Company shall not be held liable for any damages 
or delays caused by defective material, and no allowance will be made for repairs or alterations, unless 
made with its written consent or approval. 

The title and right of possession to the machinery ncrein specified, ret 
all payments hereunder, (including deferred payments and any notes or rer 
have been fully made in cash, and it is agreed that the said machinery shall i 
of the Company whatever may be the mode of its attachment to realty ( 

1 in the Company until 
s thereof, it any), shall 
n the personal property 
;e, until fully paid (or in 

cash. Upon failure to make payments, or any of them, as herein specified, the Company may retain any 
and all partial payments which have been made, as liquidated damages, and shall be entitled to take 
immediate possession of said property, and be free to enter the premises where said machinery may be 
located, and to remove the same as its property without prejudice to any further claims on account of 
damage which the Company may suffer from any cause. The company may pursue all legal remedies 
to enforce payment hereunder, but if unable to collect may thereafter repossess the property. 

The Company agrees that it shall at its own eipense defend any suits that may be instituted by 
any'party against the Purchaser, (or alleged infringement of patents relating to machinery of its own 
manufacture furnished tinder this proposal, provided such alleged infringement shall consist in the use of 
said machinery, or parts thereof, in the regular course of the Purchaser's business, and provided the 
Purchaser shall have made all payments then due under this contract, and gives to the Company 
immediate notice in writing of the institution of such suits, and permits the Company, through its 
Counsel, to defend the same, and gives all needed information, assistance and authority to enable the 
Company to do so, and thereupon in case of a final award of damage." in such suit the Company will pay 
such award, but it shall not be responsible for any compromise made without its written consent, nor 
shall it be bound to defend any suit or to pay any damages therein when the same shall arise by reason 
of the use of parts not furnished by the Company under this proposal. The Company shall also be 
notified of, and reserves the right to be represented at any tests which the Purchaser may make, iu 
relation to guarantees of operation. 

If shipment of the machinery herein specified, or any part thereof, is delayed by any cause for which 
the Company is not directly or indirectly responsible, the date of completion of said machinery by the 
Company shall be regarded as the date of shipment in determining when payments for said machinery 
are to be made, and the Company shall be enutled to receive reasonable compensation for storage; 
such storage to be at the risk of the Purchaser. If all the machinery should not be forwarded on the 
same date, pro-rata payments shall be made for partial shipments. All notes and securities given to 
the Company by Purchaser are taken by the Company, not in payment, but as evidence only of Pur- 
chaser's indebtedness. 

This contract is contingent upon strikes, fires, accidents or other delays unavoidable or beyond 
the reasonable control of the Company. The Company shall not be held responsible or liable for any 
loss, damage, detention or delay, from any cause beyond its control; and the receipt of the machinery by 
the Purchaser shall constitute acceptance of iu delivery and * waiver of any and all claims (or lost or 
damage due to any delay. 

Fig. 271. — Typical manufacturer's proposal (part I; this constitutes pages 1 and 2 of 
this particular proposal). 


PRICE: — The prut of said machintry is.. 

Twenty ..Eight Thousand Dollars, 

(*... 28,000.00 ), payable in New York, Chicago or Milwaukee Exekanie. 

TERMS.— Terms ol payment are as follows: 

60 .°/o, Cash upon jiresentation of BlU of lading _ 

20.°/o.Ca9h 50. aay 3 thereafter 

2Q .."/ft C88h_60dsy3. thereafter _ 

SHIPMENT:-The machinery herein specified will be shipped ^.^O 4ayS .. 

from the date of the receipt of th^ 

Anal information from the Purchaser, at the Company's works. 

The services of engineers, millwrights or mechanics furnished by the Compiny for ihe purpose 
of superintending the erection or operation of the machinery covered by this proposal, shall be paid 
for by the Purchaser, monthly and independent of the contract account, at the rate of Fifteen Dollars 
per eight hour day and regular overtime rates plus all traveling and hotel expenses, including all time 
the said parties are absent from the Company's works on the Purchaser's business; it being understood 
and agreed that during the term ol such service the said engineers, millwrights and mechanics shall be 
the Purchaser's employees, for whose acts the Company shall assume no responsibility. All labor and 
material required in connection with these services, will be furnished by the Purchaser. 

In the event it is elsewhere herein agreed that the Company shall erect the machinery herein sped- 
6ed, the Purchaser shall reimburse the Company for all expenses in connection with the erection of the 
machinery occasioned by delays, lack of facilities or apparatus to be furnished by the Purchaser or any 
acts for which the Company is not responsible. 

In the event the Company furnishes oil, 
under this proposal, (such as oil barrels, reels, < 
terms of this agreement, the value of such Carrie 
carriers, in good condition, to the proper receivir 
will credit the Purchaser the full amount previc 
randum and necessary shipping documents are 
ment is made, charges prepaid, within 

The Purchaser shall provide and i 
machinery herein specified, aga 
and the -Purchaser shall assume 

vire, cable or other material requiring special carriers, 
tc), the Purchaser will pay to the Company, under the' 
s in addition to the contract price. Vpon return of such 
g point, to be designated by the Company, the Company 
usly charged; provided, however, that invoice or memo- 
promptly forwarded to the Company and return ship- 

nths from the 

1 date of shii 

; of the Company adequate insurance for the 
in an amount fully protecting the Company, 
n case of failure to effect such insurance. 

All the terms and provisions of the contract between the parties hereto, are fully set out herein, 
and no agent, salesman or othei party is authorized to bind the Company by any agreement, warranty, 
statement, promise or understanding not herein expressed, and no modifications of the contract shall be 
binding on either party unless the same are in writing, accepted by the Purchaser and approved in writing 
by an Executive Officer of the Company. 



The foregolag proposal i 
this day of 


hereby accepted 

Fig. 272. — Typical manufacturer's proposal (part II; this constitutus pages 3 and 4 of 
this particular proposal). 

Sec. 293] 





These guai 


Brow^ and BlaoH Manufacturing Company. St. Louis. Mo. 

and specifications (orm part of proposal dated J*"®.. J ' 


The s.eam turbine unit- described in the following pages, when erected and properly adjusted 

the Purchaser's power house, w,ll carry true energy steady loads as given below at ._ 8Q per «,u 

power (actor and under constant operating conditions as set forth on page 6 of these specification,, 
with a consumption of dry steam not exceeding:— 

M One-half load (viz.:.375. K. W). .41.6 lbs per IC W. hour, a. 2. lbs per IC W. hour 

AtThree^uartersIoad (vi..56e.5K. W). 36.6 ..lbs. per K. W. 8..1bs. p„ K. W. hour 
A, Full load (vi.. :.m K. W.)...M...2...1bs. per IC W. hour...!'. ^ lbs per K, W. hour 

and auxiliary 
The above steam consumptions 

; include : 

■ power used by 

linals and i 

energy re- 

ies. Tlie above loads are the true electrical output at the general 

an exciter direct connected to this steam 

quired in the field tor excitation. When the proposal i 
turbine unit the steam required to drive same is included. When steam turbine is operating on low 
pressure stean, provision will be made for admitting a small amount of high pressure steam to keep 
high pressure blading cool 

Rated capacity of unit at 9Q ■ per cent powe 

Rated current per terminal 28.0... 

Normal Voltage ?3Q0- Cycles.._.e<>. 

Normal speed - - ^^ — 

Turbine to be operated condensing. 

Steam pressure at turbine high pressure throttle 

Steam pressure at turbine low pressure throttle 

Superheat in steam at turbine high pressure throttle 

Superheat in steam at turbine low pressure throttle 
Vacuum at turbine exhaust nozzle 2& 


. factor 750 K. W. Maximum. 


Phase_ _..a :; - 

revolutions per i 



780 BQ - ^^'''^ 

Excitation voltage 125 Appro: 

required with rated curre 



Diameter of H. P. steam inlet 

Diameter of L. P. steam inlet 

hundred per cent power (actor. Approximately 70 

will be required with the same current at eight per cent power factor. 

one minute: Field-.ISOQ volts; Armature... ^600 volts. 

5 inches 

exhaust nozzle ^ inches. 

. ..,11 inches. 


11 _ inches. 

urbine...l4 .inches 

Approximate overall length of unit above floor IB feet. 

Approximate additional length below floor..„ 3 

overall width of unit " 

_ feet. 

, above floor I 

Approximate ( 

•Approx. height of highest point of v 

Approximate shipping weight of unit - 

Approximate weight of heaviest piece to be handled in ei 
Approximate weight of heaviest piece to be handled afte 
Approximate amount of air required by generator per mil 

46&00 3£360 
erection 3180 


lount ol local COI 
(NOTE— U more loom !• no 
I pr^edence over Ihf fnnui t 


Kunes hanwna I 
.ith sublrtteri. 1 

FiQ. 273.— Typical manufacturer's proposal (part III; this constitutes pages 5 and 6 of 
this particular proposal). 



GENERAL DESCRIPTION— The steam turbine will be of the Iiorizonial, Allis-Chalmers reaction 
type, connecltfd lo the generator by a flexible coupling. The rotors of turbine and generator will each 
be earned in two bearings, so that either rotor may be handled separately. 

BLADING— The blading will be of the Company's patented construction, made of materials espe- 
cially adapted to resist corrosion, erosion and steam temperature specified . 

scale forming impurities, and i 

GLANDS — An adequate supply of clear water, free from ; 

steady pressure ol fifteen pounds by gauge at the glands, shall be furnished by the Pijrchaser. 

GOVERNOR — The governor will be provided with a hand operated synchronizer, arranged so that 
the mean speed ol the unit may be varied approximately three per cent above or below the normal. 

REGULATION — The variation in speed from half load to full load under ordinary operating condi- 
tions, will be approximately three per cent ; great or sudden variations ol load may cause approximate- 
ly five per cent momentary speed variation. 

SAFETY STOP — A separate safety stop governor will be supplied, which will automatically shut off 
the steam if unii reaches a predetermined speed in excess ol the normal. A lever for tripping safety 
stop by hand is conveniently located on unit. 

THROTTLE VALVE — Screw-operated high pressure and low pressure throttle valves will be pro- 
vided. Unit will be arranged so that high pressure steam will be automatically admitted m case low 
pressure steam supply is not sufficient to carry the load. Unit may be operated entirely with low pres- 
sure steam or entirely with high pressure steam. 

GAUGES — The Company's standard equipment of gauges and gauge board will be provided on unit 

HAND OF TURBINE— The turbine will be according to the Company's standard practice. 

ARMATURE — The armature core will be built up of laminated steel held in slots in the cast iron 
frame. Ventilating spaces will be provided through which air will be forced. The coils, thoroughly 
insulated, will be firmly held in slots in the laminated core. A supply of clean cool air for generator 
shall be arranged lor and supplied by Purchaser. 

FIELD — The core of the revolving field will be made of steel with slots to receive the windings. The 
windings will be ol copper securely held in the slots by wedges. The ends of the coils will be sub- 
stantially supported The alternator will be ventilated by air forced through all parts by means of fans 
attached to the field. 

are not included. 

EXCITER— The exciter is not included unless so specified. Connections to sam 
When exciter is included Purchaser shall promptly advise winding desired for ! 

RHEOSTAT— The alternator will be provided with a field regulating rheostat, arranged for installa- 
tion behind the switchboard; rheostat include; face-plate and means lor operating by hand from front 

TERMINALS — No terminals for armature leads are indnded, these leads will be arranged for solder- 
ing to the cables leading to the switchboard. No cables or wiring is included. 

PARALLEL OPERATION — ^This turbo-generator unit will operate in parallel with similar units; 
also with other units which fulfill the requirements tor parallel operation, and have a speed regulation 
similar to that of this unit 

LUBRICATION— A self-contained oiling^system will be supplied. The Purchaser shall furnish ade- 
quate clear cool water, free from acid or scale-forming impurities for oil cooler. The Purchaser shall 
provide lubricating oil of proper quality and suitable character. 

PAINTING— All exposed unfinished parts will be painted with one coat of black paint before ship- 
ping. No ornamental painting or painting after shipment is included. 

TOOLS— The Company's standard equipment of wrenches and tools will be furiiished. When more 
than one turbine is included in the contract, only one set ol wrenches and tools will be furnished. 

FOUNDATIONS— The Purchaser shall provide suitable foundations, including material and labor 
for grouting under the unit alter same has been lined up and leveled by Company's engineer, also 
such sub-loundalions, air cleanser and air ducts, for which the Company does not furnish drawings, 
as the local conditions necessitate. The foundations and sub-foundations must be so constructed that 
they will not receive or transmit vibrations from or to the adjacent flooring or structure. The Com- 
pany will furnish its standard outline and foundation plan drawings of apparatus furnished under 
these specifications. Purchaser shall furnish drawings of foundations, air ducts, etc., and shall sub- 
mit same to Company before any work is done. The Purchaser shall furnish foundation template aifd 
foundation bolts and washers. 

PIPING— The Purchaser shall furnish all steam and exhaust pipir 
ditions at turbine and shall arrange same so that no strains or vi 
turbine. The exhaust pipe must be securely anchored under exha 
construction, and must be provided with a suitable expansion joir 
Purchaser shall provide suitable size exhaust free to atmosphere provided with a water sealed auto- 
matic relief valve, if a gate valve is located in the turbine exhaust line, this atmospheric connection 
must be placed on the turbine side of same. Purchaser shall provide proper relief valve in low pres- 
sure line to turbine also proper drains and traps for all piping and shall furnish an efficient steam and 
oil separator near turbine L P. throttle ; also an efficient steam separator near turbine H. P. throttle. 
The arrangement ol all steam and exhaust piping shall be submitted by the Purchaser to the Company 
belore any work is done. Purchaser shall furnish all water piping to and from uniL 

OPERATION^Tbe steam turbine unit will operate successfully alter being properly erected and ad- 
justed, provided it receives such care and attention as is necessary and usual for units of this type 
and size; this includes the proper operation ol the condenser and ol the boiler plant, avoiding slugs 
of water and unduly wet steam also great or sudden fluctuations ol temperature or pressure. It is 
uiTderstood that the usual operating conditions will be as specified herein. 

3t ample size to give contract con- 
itions will be transmitted to the 
nozzle, laid out to avoid a stilT 
t the turbine exhaust nozzle. The 

FiQ. 274. — Typical manufacturer's proposal (part IV; this constitutes pages 7 and 8 cf 
this particular proposal). 

Sec. 294] 



294. Turbine Specifications And Guarantees (Figs. 273, 
274 and 275) are sent with the manufacturer's proposal 
(Figs. 271, and 272) and form a part of the proposal. Although 
the proposal which is here shown is for a mixed-pressure turbo- 
alternator unit, it is typical of those furnished for all classes of 
turbines. The proposal, when accepted and signed, by both 


PACKING FOR SHIPMENT— The turbine and generator will be prepared 
manner for domestic rail shipment. Packing for foreign shipment or 
included unless sp specified. 

IN GENERAL — These specifications cover the Company's standard turbine-generator unit with 
standard equipment complete as described, beginning at the mlet Hange of the throttle valves, and 
ending at the Hange of the exhaust nozzle and at the generator terminals. It is advisable that the Pur- 
chaser provide the Company promptly with drawings of the power house in the vicinity of the turbine 
location, showing other machinery, columns and foundations, existing and proposed piping, proposed 
arrangement of condensing apparatus, etc Purchaser shall provide proper space for installing unit 
and for removal of generator rotor. 

■■ readv 

desired to determine that the unit fulfills the guarantees set forth in these 
! made at Purchaser's plartt by and at the expense of Purchaser, and within 
ions. The Purchaser shall give the Company 
:e of his intention to make tests, and shall permit the Company at its expense 
in the power plant p^ior to and during tests; and to furnish and couple up such 
to the tests the Company shall have reasonable ac- 
r shall make necessary preliminary 

tests. The 

TESTS— When 

specifications sa 

thirty days aftei 

two weeks' writ 

to haverepresei 

instruments as the Company may desire. Pri' 

cess to the unit for examination and the Purch; 

ditions under which tests will be made, calibration of instruments, methods 

shall be mutually agreed upon between the Purchaser and the Company; m general the rules of the 

A. S M. E. and the A. I. E. E. will be followed. Insulation tests will be made according to the rules 

of the A. I. E. E. 

GENERATOR TEMPERATURES — Generator temperatures will be measured in accordance with 
the Standardization Rules of the A. I. E. E. as foHows: Stator: For units 500 KVA or smaller, by 
Ihcunometer applied to the hottest accessible part of the completed machine; to the temperature so 
determined will be added 15°C. correction. For units over 500 KVA the temperature will be meas- 
ured by embedded resistance coils placed as nearly as possible at the hottest part of the winding; to 
this temperature will be added 5°C. correction. Resistance temperature coils are included, but no in- 
strument will be furnished. Rotor- will be measured by increase of resistance of the winding; to the 
temperature so determined will be added 10*C. correction. 

ERECTION — For the purpose of superintending the erection and starting of the machinery described 
herein, the Purchaser agrees to and will engage and p^y for the services of such erecting engineers to 
be furnished by the Company as may be necessary, as provided in attached proposal. If, however, this 
proposal requires the Company to furnish engineers at its expense, the Purchaser shall place machin- 
ery in power house adjacent to turbine foundation and the erection of the machinery shall commence 
immediately upon engineer's arrival at Purcliaser's plant and proceed to completion without delay. 
The turbine engineer will remain at the Purchaser's plant, for operation, not longer than one week 
after the machinery is erected, it being understood and agreed that Purchaser's part of the work will 
be completed when erection of steam turbine unit is complete. The Purchaser shall pay the Company 
for the time and expenses of the engineer beyond this period ; also all time and expenses caused by de- 
lays which occur in the erection, starting, or operation of the machinery, provided the Company is not 
responsible for such delays. It is understood that the erecting engineers will not work more than ten 
hours per working day Overtime and night work also work on Sundays and Legal holidays, must be 
especially arranged for between the Purchaser and the Company. The Purchaser shall give the Com- 
pany at least one week's written notice of the date when he will be ready for the erecting engineer. 


Fig. 275.- 

- Typical manufacturer's proposal (part V; 
particular proposal). 

this constitutes page 9 of this 

the purchaser and manufacturer, forms a binding contract 
between the two. By the contract, the manufacturer can be 
held to the fulfillment of the specifications and the guarantees. 
If the turbine in an acceptance test (see Fig. 275) does not per- 
form as well as is stipulated in the guarantee, the purchaser 
has the right to reject the machine or to receive a liberal reduc- 
tion in the specified purchase price. 

295. In Selecting The Best Steam Conditions Under Which 
To Operate A Contemplated Turbine, as must be done when 


an entire plant is being designed, the unit operating cost (Sec. 
271) is again the deciding factor. By computing the unit 
operating cost for various steam conditions, those conditions 
can be found which afford the least unit cost. Generally 
speaking, the operating costs of turbines decrease (see Div. 13) 
with higher initial pressures, higher superheats, and lower back 
pressures; but the operating costs of the boilers and condensers 
go up as those of the turbines go down. Hence, the selection 
of the best operating conditions is again a matter of economics 
and must be executed with a view toward attaining the mini- 
mum unit operating cost. 

Note. — The Most Usual Steam Conditions In Turbine Plants 
are: (1) For small plants (up to about 1,000 kw.) initial pressures of 
150 to 200 lb. per sq. in. gage, superheats up to about 125° F., and vacuums 
of 27.5 to 28 in. (2) For large plants initial pressures of 200 to 300 lb. per 
sq. in. gage, superheats up to about 200° F., and vacuums of 28.5 to 29 in. 
The tendency is toward the use of higher steam pressures; several plants 
have been built for 350 lb. boiler pressure. 


1. What are the three principal objects of studying turbine operating costs? 

2. Enumerate eight factors which are usually considered as items of operating cost 
and arrange them into two groups. What are the names of the two groups? 

3. What is the usual annual amount of the fixed charges for turbines? How is the 
amount determined in any given case? 

4. Why are the fixed charges so called? 

5. Explain the meaning of the term U7iit charges. Unit operating cost. 

6. State as many factors as you can that affect the unit operating charges of a plant 
and show their effect. 

7. Define load factor and show how it affects the unit operating costs and the annual 
operating and fixed charges. 

8. What other operating costs are generally included with those of a turbine? 

9. What three factors must be considered when selecting the type of prime mover 
for a given service? 

10. For what classes of service is the steam turbine best adapted? Why? 

11. What classes of services are quite beyond the field of the steam turbine? Why? 

12. State what you can regarding the reliability of steam turbines. 

13. Upon what does the efficiency or steam economy of a turbine depend? 

14. About what steam rates may be expected from each of the following-sized turbines 
when operating condensing and when operating non-condensing: 50-kw.? 200-kw.? 
500-kw.? 1,000-kw.? 2,000-kw.? 3,500-kw.? 

15. State how the efficiency of a turbine varies with the load which it delivers. 

16. What is the meaning of a turbine's capacity rating? 

17. What can you say of the economy, in dollars and cents, of low- and mixed-pressure 
turbines? Explain. 

18. Is it advisable, usually, to employ separate high- and low-pressure turbines? 


19. Wherein does the economy of bleeder turbines he? Explain. 

20. How would you predict the steam rate of a contemplated turbine? 

21. Upon what do the relative economies of steam turbines and steam engines depend? 

22. State several advantages which, in general, the steam turbine has over the steam 
engine and vice versa. 

23. What can you say, in general, of the relative steam economies of non-condensing 
engines and turbines? 

24. In general, which has the better steam economy, a condensing engine or a condens- 
ing turbine? 

25. State the principal services for which turbines and engines of small capacity are 
each adapted. 

26. How do the prices of steam turbines vary with their capacities? Give some 
typical prices. 

27. Enumerate the steps which should be taken in selecting a prime mover for a 
given service, explaining each step as fully as possible. 

28. State briefly what information should be given to the turbine maiiufacturer when 
a quotation is requested. 

29. What is the purpose of performance specifications and quarantees in steam-turbine 
proposals? How are they enforced? 

30. How are the best steam conditions for a proposed turbine plant determined? 
What are the most usual steam conditions in practice? 


1. From the total-heat-entropy chart of Fig. 15, Hi = 1,210 B.t.u. 
H2 = 1,022 B.t.u. Hence, heat liberated = Hi - H2 = 1,210 - 1,022 = 
188 B.t.u. per lb. 

2. From the total-heat-entropy chart of Fig. 15, Hi = 1,189 B.t.u. 
H2 = 887 B.t.u. Hence, heat liberated = Hi - H2 = 1,189 - 887 = 
302 B.t.u. per lb. 

3. By For. (2): y = 223.7VHi - H2 = 223.7Vl88 = 223.7 X 13.7 = 
3,065 ft. per sec. Or v = 3,065 X 60 ^ 5,280 = 34.8 mi. per min. By 
charts B and C, Fig. 15, v = 3,050 ft. per sec. 

4. By For. (17): W = 778W(Hi - H2) = 778 X 1 X 302 = 235,000 

5. By For. (18): P = W(Hi - H2)/2,545 = 2,000 X 302 -^ 2,545 = 
237 hp. By For. (19): Wh = 2,545/(Hi - H2) = 2,545 ^ 302 = 8.43 
lb. per hp.-hr. From AB, Fig. 15: Wh = 8.4 lb. per hp.-hr. 

6. From Fig. 20, for a 2,000-hp. turbine: E^ = 65 per cent = 0.65. 
Hence, by For. (26): Wb = 2,545/[Er X (Hi - H2)] = 2,545 - [0.65 X 
188] = 20.8 lb. per b.hp-hr. Hence, at full load, W = Pb X Wb = 
2,000 X 2a8 = 41,600 lb. per hr. 

7. By Sec. 16, for maximum work: Vb = t',/2 = 3,065 ^ 2 = 1,532ft. 
per sec. 


1. By For. (28): Wwi = tWsi(Li + L2) 72(^1 - T2). Now from 
steam tables, Ti = 240.1° F. T2 = 233.1° F. Li = 952. L2 = 956.7. 


Hence, Wpri = 3 X (1,500 X 25 -^ 60) X (952 + 956.7) -^ [2 X (240.1 
- 233.1)] = 3 X 625 X 1,908.7 ^ 14 = 262,772 Ih. Also, by For. (29): 
Wtf2 = Ws2(Lx + U)/2{T, - T2) = 2,000 X (952 + 956.7) ^ [2 X 
(240.1 - 233.1)] = 2,000 X 1,908.7 -^ 14 = 272,671 lb. 

2. The condenser must handle 1,500 X 25 = 37,500 lb. of steam per 
hr. By Sec. 173, the steam temperature should be at least 60 + 25 = 
85° F. From Fig. 184, the temperature at 1.5 in. pressure is 92° F. 
Hence, it is feasible to operate with this condenser pressure. The dis- 
charge circulating water temperature should not exceed 92 — 10 = 82° F. 
Assume a 20° F. rise through the condenser. The rate of heat transfer 
with this pressure may be assumed at 350 B.t.u. per sq. ft. per hr. per 
degree difference. Hence, using Fig. 184, and beginning at 1.5 in. pres- 
sure on the lower scale and following upward to the 60° F. line, to the 
left to the 20° F. rise line, upward to the 350 B.t.u. line, to the left to the 
curve and upward to the surface scale, there results a value of 125 sq.ft. 
per 1,000 lb. steam. The condenser surface = 37.5 X 125 = 4,687.5 sg. 
ft. The circulating water required = 37.5 X 95 = 3,562.5 gal. per min. 



1. By Fig. 20, the efficiency ratio = 0.55 = E^. The low-pressure steam 
rate = Wb2 = 6,000 ^ 500 = 12 lb. per hp.-hr. From the total-heat- 
entropy chart of Fig. 15, Hi = 1,225 B.t.u. per lb. H2 = 877. H/ = 
1,156. H2' = 952. Hence, by For. (30): 

^^^ = Hi - H. [^ - W^=^(H/ - H2') ] = [1 - (1.225 - 877)] X 
{(2,545 -h 0.55) - [12 X (1,156 - 952)]} = (1 4- 348) X [4,630 - (12 X 
204)] = (4,630 - 2,448) -^ 348 = 2,182 ^ 348 = 6.3 lb. per hp.-hr. 
Hence, total high-pressure steam required = 500 X 6.3 = 3,150 lb. per hr. 

2. By Fig. 20, the efficiency ratio = Er = 0.63. The extraction rate = 
Wb2 = 22,500 ^ 1,500 = 15 lb. per hp.-hr. From Fig. 15, Hi = 1,257 
B.t.u. per lb. H2 = 868. Hi' = 1,257. H2' = 1,091. Hence, by For. 

^^^ = Hi-H2 [^7^ ~ ^^^(H^' - ^''^ ]^b^ ^^'^^^ ~ ^^^^] ^ 
((2,545 - 0.63) - [15' X (1,257 - 1,091)]} =- (1 -^ 389) X [4,040 - 
(15 X 166)] = (4,040 - 2,490) ^ 389 = 1,550 -^ 389 = 3.99 lb. per hp.- 
hr. Hence, total steam required = 1,500 X 3.99 + 22,500 = 5,980 + 
22,500 = 28,480^6. per hr. 


Acceptance test, water rate correction, 295 
Accumulator, see Regenerator . 
Adjustment, see also Clearance, Align- 

axial, rotor, 100 

speed, see Governor. 
Air leak, 244 

Air-pressure governor, Ridgway, 148 
Alberger Pump & Condenser Co., 69 
Alignment, see also Clearance, Adjust- 

axial, bucket-wheel turbine, 82 

checking, 84 

coupling, 171 
Allis-Chalmers Mfg. Co., bearing, 89 

claw-type flexible couphng, 169 

correction graphs, 301 

fixed blades, 25 

gland, 105 

governor, 135 

half-cyhnder, 27 

lubrication system, 222 

oil cup on valve stem, 233 

spindle, 87 

turbine. Parsons type, 25 
single-flow reaction type, 63 
Alternator, load shifting, 244 

starting, 245 

stopping, 246 
"American Electricians' Handbook" T. 

Croft, on load division, 153 
Ammeter for turbine test, 272 
Amy, E. V., in "Electrical World" on 

vibration, 238 
Ashton Valve Co. relief valve, 156 
"A. S. M. E. Test Code" on leakage 
tests, 261 

performance values, 266 

testing, 252 

turbine test data form, 271 
Atmospheric-relief VALVE, 156 

Schutte, 182 
Auxiliary oil pump, Allis-Chalmers, 222 
Available heat, 266 
Axial adjustment, see Adjustment. 
Axial-flow turbine, adjustment, 101 

definition, 38 

single-stage, re-entry, 47 


Backlash, reduction-gears, 166 
Back pressure, decreases thermal eflB- 
ciency, 294 
increases water rate, 294 
turbine, see also Non-condensing turbine, 

water-rate effect, 295 
Balance, load, engine and turbine, 216 
Balancing load, see Heat balance. Load. 
Ball bearing, see Bearing. 
Barometer for turbine test, 272 
Barrel, definition, 27 

Barstow, J. S., on applicabiUty of tur- 
bines and engines, 320 

Base pressure and superheat, 305 
Batch treatment, oil, 225 
Bearing, bearings, 87-102 
alignment, 81 
BALL, lubrication, 220-222 

use, 92 
classification, 88 
flexible, action, 61 
operation, 90 
tubular, 91 
lubrication, circulation oiling, 221 
main, care, 93 
OIL, coohng, 229 
functions, 218 
properties, 224 
temperature, desirable, 230 
low, maintaining, 92 
plain flexible, 88 
radial pressure, 88 
repair, 94 
rigid, 91 
ring-oiled, 233 
Sturtevant turbine, 219 
thrust, adjustment, 102 
function, 95 
Kingsbury, 98 
multi-collar, 97 
roller, 96 
simple collar, 97 
types, 96 
types, 87 
water-cooled, 93 
wear, 94 
Bedplate alignment, 83 
Biddle, J. G. Co., reed tachometer, 261 
Blades, 24 

Blading, impulse, reaction, difference, 33 
impulse turbine, 31 
reaction turbine, 32 
relation to cylindrical section, 33 
repair, 248 
Blast governor, steam-relay, 134 
Bleeder diaphragm, Terry turbine, 195 
mechanism. General Electric Co., 195 

Terry turbine, 194 
turbine, 186-217 

control, steam distribution, 192 

cost, 206 

definition, 35 

economy, calculation, 206 

reasons, 315 
governing, 192 
heat balance system, 203 
speed regulation, 206 
starting, 243 
STEAM CONSUMPTION, calculation, 208 

chart, 210 
use, 190 
VALVE, General Electric Co., 193 
Terry Turbine, 194 
Blower turbine, 44 
Boiler feed water, measurement, 259 

weighing, 260 
Brake output, 265 
Branca's impulse turbine, 2 
Bucket, 24 




Bucket-wheel turbine, 24 

definition, 38 

illustration, 28 
Bypass governor, see Governor, 

valve, see Valve. 

Capacity, generating, eflfect of load factor, 

how increased, 285 
Carbon gland, see Gland. 

ring, 114 
Carling turbine-driven fan, 54 
Case, casing, 26 

protection, piping strains, 83 
"Central Stations" T. Croft on load and 

demand factors, 311 
Centrifugal governor, see Governor, 

pump, starting, stopping, 246 

water-packed gland, see Gland. 
Charges, fixed, 307 

operating, 309 
Circular section, 34 
Circulation oiling systems, 230 

Allis-Chalmers Mfg. Co., 222 
Circumferential section, 34 
Claw coupling, 169 

Clearance, see also Adjustment, Align- 

axial, checking, 82 

blade, testing, 81 

carbon gland, 114 

reduction-gear teeth, 166 

rotor, axial adjustment, 99 
Cochrane constant-pressure valve, 200 
Coil, cooling, installation, 81 
Composite-staged TrRsiNE, construc- 
tion, 60 

definition, 38 

Kerr, 62 

Moore, 59 
Composite staging, 60 
Compound turbine, 66 
Condensate, weighing, 258 
Condenser, CONDENSERS, 177-184 

connection to turbine, 184 

cooling water, formula, 179 

definition, 177 

graph, surface, water volume, 181 

installation, 180 

jet, 182 

pressure, determination, 255 

surface, formula, 179 

turbines which require, 188 

vacuum breaker, purpose, 184 
Condensing engines, water-rate, 325 


economics vs. engine, 319 
high-pressure, use, 190 
piping, 242 
starting, 241 
vacuum, 244 
water rates, graph, 325 
Continuous bypass oil treatment, 226 
Cooler, oil, cleaning, 233 
Cooling coil, installation, 81 

water, formula, 179 
CoppiTS engineering & equipment Co., 
turbine construction, 47 
turbine, illustration, 45 
turbo-blower, lubrication, 220 
Correction graphs, see Graphs. 
performance, 295 
pressure, application, 302 
Correction, test, base pressure and super- 
heat, 305 

Cost, operating, 307 

inclusion with boiler cost, 311 

load factor efifect, 310 
Cost, turbines, graph, 321 

turbo-generator, table, 313 

unit operating, 308 
Coupling, couplings, 160-172 

aligning shafts, 170 


Sturtevant, 168 
care, 170 

claw, pin, and jaw types, 169 

"flexibihty," 170 

purpose, 168 

rubber-bushing type, 168 
RIGID, 167 
Critical speed, 87 
Croft, T. in: 
"American Electricians' Handbook" on 

load division, 153 
"Central Stations" on load and demand 

factor, 311 
"Machinery Foundations And Erec- 
tion " on alignment and leveling, 77 

bedplate leveling, 78 

condenser supports, 184 

planning turbine foundation, 74 
"Practical Heat" on entropy, 11 

heat and work, forms of energy, 1 

kinetic energy, 9 

perfect engine, 257 

Rankine-cycle efficiency, 279 

steam liberating heat, 8 

temperature, 253 
"Steam-engine Principles And Prac- 
tice" on dash-pots, 152 

governors, 84 

leaf-spring governors, 125 

lubrication, 218 

measuring output, generators, 256 

oil filters, 227 

operating costs, 307 

rebabitting bearings, 94 

viscosity, 224 
"Steam Power Plant Auxiliaries And 
Accessories" on condensers, 179 

high turbine vacuum, 292 

valves, 156 
Cross-compound turbine, 67 
Curtis nozzle, 24 
stage, 60 
turbine, see also General Electric Co. 

checking clearance, 82 

illustration, 43 


definition, 44 
Cylinder, definition, 26 

half, illustration, 27 
Cylindrical section, 34 

Dash-pot, governor, 151 

Data form, turbine test, 271 

De Laval Separator Co., oil purifier, 228 

turbine governor, 119 
De Laval Steam Turbine Co., governor, 
vacuum breaker, 126 

heat-balance system, 203 

marine turbine with reduction gears, 161 

nozzle, 23 

nozzle and valve, 156 

pressure-staged turbine, 56 

shaft, 86 

single-stage turbine, 5, 46 
De Laval turbine type, 43 



Diaphragm, 56 

Direct-current generator, starting, 
stopping, 246 
Disc, 25 
Double reduction gears, see Reduction 

Double-flow turbine, 40 
Drains, installation, 83 
Drum, 25 
Dynamometer, turbine test, 272 


Economics, 307-334 
Economy, bleeder turbine, 206 

comparison, 305 

mixed-pressure turbine, 192 

relative, engine and turbine, 316 
Eddy losses, 17 
Efficiency, calculation, 268 

values, 314 
"Electrical World" E. V. Amy on vibra- 
tion, 238 
Emergency governor, see Governor. 

valve, see Valve, safety stop. 
Emulsification, oil, 224 
Energy losses, 17 
Entropy chart, steam, 10 

definition, 11 
Energy, heat, see Heat energy. 

kinetic, see Kinetic energy. 
Engine and turbine, floor space, 317 

head room, 317 

load, balance, 216 

relative economy, 316 
Engine, heat, 1 

steam, advantages, disadvantages, 317 
Epicyclic gear, see Reduction gears. 
Erection, turbo-generators, 78 
Exhaust steam, balance, 202 

properties, dettirmination, 255 

superheated, loss, 288 

velocity loss, 17 
Exhaust-steam turbine, see Low-pressura 

Expansion joint, low-pressure, 182 

Westinghouse rubber, 183 
Extraction turbine, see Bleeder turbine. 

Feed water, boiler, measurement, 259 

Fixed blades, 25 

Fixed charge, see Charge. 

Flexible coupling, see also Coupling. 

purpose, 168 
Float-valve water-level control, 174 
Floor space, engine and turbine, 317 
Flow valve, use, 199 
Fluid stream, forces produced, 3 

velocity reduction, 12 
Force-feed lubrication, see Circulation oil- 
Forces due to fluid streams, 4 
Foundations, turbine, 75 
Frictional losses, 17 
Fuel saving due to superheat, 290 

Gages, turbine test, 272 

Gear, reduction, see Reduction gear. 

Geared turbine, lubrication, 230 

Gebhardt "Steam Power Plant Engineer- 
ing " on operating-charge comparison, 

General Electric Co. bearing, 90 
bleeder valve, 193 
carbon gland, 112 
composite-staged turbine, 60 
governor, centrifugal, 146 

inertia, 148 

inertia and centrifugal, 147 

lubrication, 154 

multi-valve, 145 

valve, multi-ported, 144 
installation, 81 

marine turbine oiling system, 210 
on checking alignment, 84 

receiving and unpacking, 75 

spring-opposed governors, 145 
Rice mechanical valve gear, 150 
synchronizing device, 152 
throttle valve, 155 

velocity-and-pressure-staged turbine, 58 
water-cooled bearing, 93 
"General Electric Review" on steam con- 
ditions, turbine, 288 
Generator, direct-current, starting and 

stopping, 246 
output, determination, 256 

thermal efficiency, 265 
Gershberg, Joseph, in "Power" on steam 

consumption of bleeder turbine, 209 
Gland, 103-117 
carbon-packed. 111 
centrifugal water-packed, 108 
definition, 27, 103 
labyrinth, 105 
metallic-packed, 103 
steam-seal leakage, 114 
types, 103 
Governor, 119-154 

adjustment in synchronizing, 152 

Westinghouse, 139 
air-pressure, Ridgway turbine, 148 
bleeder turbine, 192 
care, 153 

centrifugal-and-inertia, 147 
centrifugal, direct throttling, 123 
cent-rifugal. General Electric Co., 146 

oil-relay intermittent, Westinghouse, 
classification, 123 
Curtis turbine, lubrication, 154 
De Laval oil-purifier turbine, 119 
direct, 27 

DIRECT throttling, 120 

adjustment and troubles, 126 
emergency, adjustment, 132 

definition, 27 

illustrations, 128-131, 142 

maintainance, 132 

testing, method, 131 
frequency, 154 
function and operation, 120 
hunting prevention, 151 
indirect, 27 
inertia, 148 

inspection after installation, 84 
intermittent, 134 

Westinghouse, 138 
mechanical indirect, 150 
mechanism, 121 
mixed-pressure turbine, 197 
oiling, 155 
oil-pressure, 122 
oil-relay, throttling, 133 
relay, definition, 27, 122 
spring-opposed, adjustment, 145 
steam-pressure, 122 
Sturtevant, adjustment, 124 
throttling, direct centrifugal, 123 



Governor, vacuum breaker, 125 
VALVE, definition, 155 

low-pressure turbine, transverse sec- 
tion, 216 
mixed-pressure turbine, 199 
Terry turbine, 127 
Westinghouse centrifugal, adjustment, 
Graphs, water rate correction, pressure, 

superheat, vacuum change, 300 
Gravity circulation, see Circulation oiling. 
Guarantees and specifications, 335 
Guide blades, 25 
Gurney thrust bearing, 98 

Jet impulse effect, 3 
impulsive, 18 
reaction, 19 


Kerr Turbine Co., turbine, 62 

lubricating system, 220 

oil cooler, 230 

sentinel valve, 157 
Kerr tubo-generator bearing, 90 
Kinetic energy, acquired by steam, 8 

work conversion, 14 
Kingsbury thrust bearing, 98-99 


Head room, engine and turbine, 317 
Heat and load, available, 267 

balance, 202 

consumption, turbine, 270 

conversion, perfect turbine, 15 

energy chart, 315 

conversion into work, 2 
relation to kinetic, 9 

engine, 1 

input, total, 267 

liberated by steam, 8 
Heat-entropy chart, steam, 10 
Hero's turbine, 1 
High-pressure turbine, definition, 34 

diagram, 205 

uses, 189 
Horizontal turbine, 43 
Horsepower, commercial turbine, 17 

perfect turbine, 15 
Hunting, 151 
Housing, 26 

Impulse blading, 33 
Impulse, definition, 2 

turbine and reaction, differences, 32 

Branca's, 2 

characteristics, 31 

composite-staged, 60 

definition, 30 

glands, 107 

pressure-staged, 55 

single-stage, 45 

staging, terminology, 37 

types, 43 

velocity staged, 52 

velocity-and-pressure staged, 58 
Impulse-and-reaction turbine, con- 
struction, 67 
definition, 34 
Impulsive force, 3 

jet, 18 
Inertia governor, 148 

Ingersoll-Rand Co., composite-staged tur- 
bine, 61 
Input, heat, definition, 266 
Inspection, turbine, 247 
Installation, 74-85 
condenser, 180 
Curtis turbine, 81 
procedure, 74 
Insulating shims, 81 
Instruments, turbine test, 272 
Intermittent governor, 134 

Jaw coupling, 169 
Jet condenser, 177 

Labyrinth gland, see Gland. 
Leakage losses, 17 
Ljungstrom turbine, 39 
Load, alternator, shifting, 244 

and heat balance, 202 

application, starting turbine, 245 

balance, engine and low-pressure tur- 
bine, 213 

factor, determination, 323 
generating capacity effect, 311 
operating cost effect, 310 

fractional, efl5ciency, 314 

shifting, 244 

steam consumption, 271 

taking off, 246 

turbine, insufficient power, 249 
Losses, causes, 283 

energy, 17 
Low-pressure turbine, 186-217 

cost, 212 

definition, 34 

flow valve, 199 

function, 209 

governor, 216 

load balance, 213 

piping, 213 

regenerator, 173 

steam economy, 314 

uses, 211 
Lubrication, see also Oil, Oiling, 218-234 

ball bearings, 220 

circulation oiling, 221 

geared turbine, 230 

governor, Curtis turbine, 154 

Kerr turbine, 220 

oil, 224 

system, care and operation, 231 


"Machinery Foundations And Erec- 
tion" T. Croft on alignment and 
leveling, 77 
condenser supports, 184 
expansion joints, 184 
planning turbine foundation, 74 
Main bearing, see also Bearifig, 87 
Maintenance, 247 
Manufacturer's proposal, 331-335 
Manufacturers, turbines, table, 69-71 
Marine turbine reduction gears, 161 
Marks, "Mechanical Engineers' Hand- 
book" on water rates, 313 
Marks and Davis, "Tables and Diagrams 
of The Thermal Properties of Satu- 
rated and Superheated Steam," 12 
Metal packing rings, 104 
MetalUc-packed gland, see Gland. 
Midwest Engine Company, turbine, 48-50 
Mixed-flow turbine, see Mixed-pressure 



Mixed-pressure turbine, construction, 
cost, 206 
definition, 35 
economy, 206 
flow valve, 199 
governor, 197 
illustration, 196, 202 
speed regulation, 206 
starting, 243 
steam consumption, 208 
steam economy, 314 
uses, 201 
Moore Steam Turbine Corp., composite- 
staged turbine, 59 
construction, 55 
GOVERNOR, direct, 121 
emergency, 130 
relay, 132 
reduction gears, 160 
velocity-staged turbine, 30 
Moving blades, 24 

Moyer, J. A., in "Steam Turbines" on 
definition of "impulse" and "reac- 
tion," 3 
Multi-ported governor valve, 144 


National Electric Light Association 
"Prime Movers Committee's Report" 
on selecting prime movers, 322 
Net output, generator, 257 
Non-condensing turbine, definition, 34 

economy relative to engine, 318 

high-pressure, use, 189 

piping, 240 

plant diagram, 205 

starting, 240 
Nozzle, definition, 23 

De Laval, 156 

fouled, 244 

moving, maximum work, 19 

shape, effect on velocity, 11 

steam action in, 8 

Sturtevant, 51 

Terry, 53 

valve, see Valve. 

Oil, breaking down, 232 

cooler, cleaning, 233 
construction, 229 

corrosive, 224 

emulsification, 224 

filters, 227 

function in bearing, 218 

impurity deposits, causes, 228 

level, ring-oiled bearing, 233 

manufacturers' recommendations, 225 

method of supplying, 218-221 

pump, auxiliary, AUis-Chalmers Mfg. 
Co., 222 

properties, 224 

purification, 225 

shield, 234 

temperatures, 92, 230 

treatment, 225 

viscosity, 224 
Oil-relay governor, see Governor. 
Oiling, see also Lubrication. 

circulation, see Circulation oiling. 

gravity system, 219 

ring, 218-221 

system, Ridgway turbine, 223 

Operation and maintenance, 235-250 

fundamentals, 235 

general precautions, 235 

safety rules, 236 

steam conditions, 284 

troubles, 237 
Operating charge, see Charge. 

cost, see Cost. 

faults, location by test, 271 
Output, power, determination, 255 
Overspeed governor, see Governor, emer- 

Overload valve, 157 

Packing gland, see Gland. 

ring, see Ring. 
Parsons, as turbine developer, 2 
Parsons Marine Steam Turbine Co., tur- 
bine and reduction gears, 162 
Parsons turbine, see Reaction turbine. 
Pelton water wheel, 4 
Performance, comparison, 275 

values, formulas, 265 
terms, 261 
Pin couphng, 169 

Piping, centrifugal water-packed eland, 

condensing turbine, 242 

lubricating system, 220 

layout, testing, 258 

non-condensing turbine, 240 

precautions, 82 

regenerator accessories, 174 

steam-seal, 112 

strains, protection, 83 

turbine, 74 
Pitot tube. 244 
Plain bearing, 88 

"Power" E. H. Thompson on erection, 

fitting carbon ring, 116 

J. Gershberg, on steam consumption of 
bleeder turbine, 209 

on exhaust-steam heat balance, 204 

on steam conditions, Westinghouse tur- 
bine, 280 
Power output, determination, 255 

plant, heat balance, 201 
Power "Turbo-Blower Co., turbine, 44 
"Practical Heat," Croft, T. on 
entropy, 11 

forms of energy, 1 

kinetic energy, 9 

on Rankine-cycle efficiency, 279 

perfect engine, 267 

steam liberating heat, 8 

temperature reading, 253 
Pressure change, condenser, determina- 
tion, 255 correction graph, 300 

operating, regenerator, 175 

stage, definition, 37 

STAGING, definition, 19 
purpose, 57 

STEAM, advisable, 287 
effect of change, 286 
governor system, 122 
Pressure-staged turbine, definition, 37 

hydrauhc, 20 
Poole Engineering and Machine Co. 

reducing gears, 166 
Prime-mover selection, factors, 311 

procedure, 322 
Proposal, turbine, 331-335 
Pump, centrifugal, operation, 246 



Quotation, requesting, 330 

Radial-flow turbine, 38 
Radial-pressure bearing, 88 
Radiation losses, 17 
Rankine-cycle efficiency, 268 
RATIO, as performance value, 262 
determination, 265 
significance, 268 
Rateau nozzle, 24 
regenerator, 173 
stage, 60 
turbine, 43 
Rating of turbines, 314 
Reaction, definition, 2 
jet, 19 
stage, 37 
staging, 21 

turbine, and impulse, blading, 32 
characteristics, 23 
cross-compound, 66 
definition, 31 
differences, 32 
double-flow, 65 
forces, 7 
glands, 106 
half cyhnder, 27 
Hero's, 1 

operation explained, 63 
radial-flow, 39 
single-and-double-flow, 65 
single-flow, 64 
tandem-compound, 66 
types, 64 
Reactive force, 5 
Reducing valve, use, 199 
Reduction gears, 160-172 
alignment, 164 
classification, 161 
construction, 163 
definition, 29 
efficiency, 164 
epicyclic, 166 
function, 160 
lubrication, 164 
purpose, 161 
tooth clearance, 166 
troubles, 164 
uses, 161 
Re-entry type, definition, 45 
Regenerator, 173-177 
definition, 173 
formula, 176 
operating pressure, 175 
piping accessories, 174 
practicability, 175 
Rateau, 173 
Regulation, speed, 121 
Relay governor, 27, 122 
Relief VALVE, Ashton, 156 
function, 156 
Schutte, 182 
Sturtevant, 26 
Repulsive force, definition, 5 
Reversing chamber, axial-flow turbine, 
buckets, tangential-flow turbine, 51 
Rice mechanical valve gear, 150 
Ridgway Dynamo & Engine Co., gover- 
nor, 148 
high-pressure turbine, 57 

clearances, 56 
oiling system, 223 
Rigid coupUng, see Coupling. 

Ring, carbon, refitting, 115 

oihng, 218-221 

packing, metal, 104 
Rotor, see also Shafts Spindle. 

assembled, rigid coupling, 168 

axial adjustment, 99 

definition, 25 

reaction turbine, 87 

speed determination, 261 
Runner, 25 


Safety stop, see Governor, emergency. 
Safety-stop valve, see Valve. 
Schutte & Koerting automatic flow 
regulating valve, 201 
free exhaust valve, 182 
jet condenser, 179 
Seal, steam, operation, 107 

piping, 112 
Section, cyhndrical, 34 
Selection, prime mover, 322 

turbine, 307-334 
Semi-double-flow turbine, 40 
Sentinel valve, 157 
Shaft, see also Rotor, Spindle. 
ahgning at coupling, 170 
construction, 86 
critical speed, 87 
definition, 25 
flexible, 87 
stiff, 87 
Shims, axial adjustment, 102 

insulating, 81 
Single-and-double-flow turbine, 40 
Single-entry turbine, 44 
Single-flow turbine, 39 
Single reduction gear, see Reduction gear. 
Single-stage turbine, 37 
Sludge, 225 
Specifications, 335 
Speed, adjustment, see Governor. 
control by governor, 120 
critical, 87 

governor, see Governor. 
reducer. Turbo-gear, 166-167 
regulation, bleeder turbine, 206 
formula, 121 

mixed-pressure turbine, 206 
Spindle, see also Shaft, Rotor. 

definition, 86 
Stage, definition, 35 

valve, see Valve, bypass. 
Staging, definition, 19 
impulse turbine, 37 
pressure, 57 
Stationary blades, 25 
Stator, 26 

Steam, action in turbine, 2 
chest, 122 
CONDITIONS, determination, 252 

EFFECT ON thermal efficiency, table, 
water rate, 285 
selection, 335 

table by manufacturers, 70 
turbines, for different, 186 
Westinghouse turbine, 280 
CONSUMPTION, bleeder turbine, 208 
determination, 257 
graph, 263 
high, causes, 248 
metering, 261 

mixed-pressure, turbine, 208 
various loads, 271 
distribution, bleeder turbine control, 192 
economy, 314 



Steam engine, see Engine. 
"Steam-engine Principles And Prac- 
tice BY T. Croft, on dash-pots, 152 

governors, 84 

leaf-spring governors, 125 

lubrication, 218 

measuring generator output, 256 

oil filters, 227 

operating costs, 307 

rebabbitting bearings, 94 

viscosity, 224 
Steam, expansion in nozzle, 8 

exhaust, see Exhaust. 

heat-entropy chart, 10 
Steam Motors Co. two-bearing tur- 
bine, 55 

assembled rotor, 168 

casing, 55 

emergency governor, 131 

gland, 106 

governor, 128 
"Steam Power Plant Auxiliaries And 
Accessories" by T. Croft on con- 
densers, 179 

turbine vacuum, 292 

valves, 156 
"Steam Power Plant Engineering" by 
Gebhardt, on operating-charge com- 
parison, 310 
Steam pressure, see Pressure. 

rate, turbine, 316 

reaction wheel, 1 

relay governor, see Governor. 

seal, see Seal. 

strainers, 157 

superheated economy, 288 • 

temperature, determination, 255 

turbine, see Turbine. 
"Steam Turbines" by Moyer on definition 

of "impulse" and "reaction," 3 
Steam, velocity, 9 
Steam-sealed gland, see Gland. 
Strainer, installation, 83 

purpose, 157 
Stuffing-box gland, 103 
Sturtevant, B. F. Co., turbine, bearing, 

coupling alignment, 168 

emergency governor, 128 

exterior view, 52 

governor adjustment, 124 

main governor, 123 

nozzle and reversing buckets, 51 

relief valve, 26 

section, 28 
Superheat, advisable, 289 

change, water-rate correction graph, 300 

effect, 278-306 

fuel saving, 290 

increase, eflFect, 284 
Supply-steam pressure, increase, effect, 

Surface condenser, 177 
Sweetening oil treatment, 226 
Synchronizing, governor adjustment, 152 

Tachometer, electric, 262 

for turbine test, 272 

vibrating-reed, 261 
Tandem-compound turbine, 66 
Tangential-flow turbine, axial adjust- 
ment, 101 

definition, 38 

single stage, re-entry, 51 
Tanks for turbine test, 272 

Terry Steam Turbine Co., bleeder 
mechanism, 194 
carbon-ring gland, 113 
composite-staged turbine, 61 

blade clearances, 53 
emergency governor, 131 
governor, 127 
tangential-flow turbine, 51 

lap, 100 
mixed-pressure turbine, governor dia- 
gram, 198 
ON steam consumption, 248 
turbine load, 248 
writing for advice, 249 
Test, acceptance, water rate correction, 
correction, values, 305 
turbine, data form, 271 
Testing, 251-276 

apparatus and instruments, 252 
data required, 251 
duration of tests, 252 
log sheet form, 264 
purpose, 251 
Thermal efficiency, as performance 
value, 262 
dependent conditions, 278-306 
decreased by back pressure, 294 
effect of steam conditions, 279 
generator output, 265 
significance, 269 
Thermometers for turbine test, 272 
Thompson, E. H., on erection, 78-80 

fitting carbon ring, 116 
Throttle valve, see Valve, 143 
Throttling governor, see Governor. 
Thrust bearing, see also Bearing, 88 

Kingsbury, 96 
Total heat input, 266 

Troubles, operating, location by test, 271 
Tube, condenser, fouled, 244 
Turbine, adaptabiUty, 312 

advantages and disadvantages, 317 
ahgnment, 81 

and engine, applicability, 320 
floor space, 317 
head room, 317 
approximate horsepower and water rate, 

axial-flow, see also Axial-flow turbine. 
back-pressure, 34 
bearing, see Bearing. 
bleeder, see Bleeder turbine. 
Branca's, 2 
bucket-wheel, definition, 38 

illustration, 28 
capacity, how to increase, 285 

rating, 314 
care while running, 243 
classification, 23-41 

table, 29, 42 
composite-staged, 38 
compound, 66 
condenser, see Condenser. 
condensing, see Condensing turbine. 
cost, graph, 321 

table, 313 
coupling, see Coupling. 
Curtis, 59 
double-flow, 40 
economics, 307 

ECONOMY, effect of steam conditions, 
relative to engine economy, 316 
efficiency, 314 
efficiency ratio, 17 
energy losses, 17 



Turbine, equal-pressure, 30 

exhaust-steam, see Low-pressure turbine. 

extraction, see Bleeder turbine. 

foundations, 75 

geared, lubrication, 230 

gland, see Gland. 

governor, see Governor. 

heat consumption, 270 

Hero's, 1 

high-pressure, see High pressure turbine. 

history, 1 

horizontal, 43 

hydraulic, pressure-staged, 20 

IMPULSE, see Impulse turbine. 

and reaction, diflferences, 32 
inspection, 247 
installation, see Installation. 
LOAD balance, 201 

insufficient power, 248 
low-pressure, see Low-pressure turbine. 
lubrication, see Lubrication. 
maintenance, 247 

maximum economy, operation, 237 
mixed-flow, see Mixed-pressure turbine. 
mixed-pressure, see Mixed-pressure tur- 
manufacturers, table, 69-71 
nomenclature, 23-41 
non-condensing see Non-condensing tur- 
bine, 34, 189 
nozzle, see Nozzle. 
operation, see Operation. 
Parsons, see Reaction turbine. 
PERFOKMANCE, Comparison, 275 

values, terms and efficiencies, 261 
Piping for small, 74 

precautions, 82 
placing on foundation, 76 
power output, determination, 255 
pressure-staged, 37 
principles, 1-22 
proposal, 331-335 
quotation, requesting, 330 
radial-flow, 38 

reaction, see Reaction turbine. 
receiving and unpacking, 76 
reduction gear, see Reduction gear. 
regenerator, see Regenerator. 
reliability, 312 
reversibility, 312 
rigid-coupHng, two-bearing, 167 
rotor, see Rotor. 
selecting steam conditions, 335 
selection, 307-334 
semi-double-flow, 40 
shaft, see Shaft. 
single-and-double-flow, 40 
single-flow, 39 
single-stage, 37 

specifications and guarantees, 335 
speed, see Speed. 
stage, 35 
starting, 239 
STEAM conditions, 186 

consumption, 248 

economy, 314 
stopping, 245 
tangential-flow, see also Tangential- flow 

testing, see Testing. 
types and construction, 42-73 
unequal-pressure, 31 
usual steam conditions, 334 
valve, see Valve. 
velocity, 30 

velocity- and pressure-staged, 38 
VELOCITY-STAGED, application, 54 

Turbine, telocitt-staged, definition, 37 

vertical, 43 

water rates, 313 
Turbine-room, log sheet, 243 
Turbo-alternator, power output, 256 
Turbo-gear speed reducer, 166-167 
Turbo-generator, cost, 313 

power output determination, 256 

sets, erection, 78 

water rates, 313 


Unit cost, see Cost. 

Vacuum breaker, governor-operated, 126 

in condenser, 182 

CHANGE, effect on water rate, 293 
water-rate correction graph, 300 

effect, 278-306 

maintaining, 244 

usual, turbine practice, 292 
Valves, 155-158 

flow, see Flow valve. 

free exhaust, 182 

gear, Rice mechanical, 150 

governor, see Governor. 

nozzle, 51, 156 

reducing, see Flow valve. 

reUef, 26 

throttle, Westinghouse, 143 
Vanes, 24 
Velocity, acquired by steam, 9 

energy, 2 

moving nozzle, maximum work, 19 

stage, 36 

staging, 20 

steam, effect of change, 284 
Velocity- and pressure-staged turbine, 38 
Velocity-staged turbine, 37 
Vertical turbine, 43 
Vibration, 238 
Viscosity, oil, 224 
Voltmeter for turbine test, 272 


Water, condenser, determination, 181 
Water-packed gland, see Gland. 
Water rate, as performance value, 266 
approximate formula, 17 
condensing turbines and engines, graph, 

CORRECTION of test valucs, 295 

graph, pressure, superheat, vacuum 

change, 300 
EFFECT OF back pressure change, 295 
steam conditions, table, 279 
superheat change, 291 
supply-steam pressure change, 287 
vacuum change, 293 
formula, 264 
graph, 263 
perfect turbine, 15 
turbo-generator, table, 313 
various loads, 271 
Water, regenerator formula, 176 
Waterwheel, Pelton, 4 
Wattmeter for turbine test, 272 
Wedges, for turbine alignment, 76 
Westinghouse Electric & Mfg. Co., 
automatic throttle valve, 143 
bearing, 89 



WestinGhottse Electric & Mfg. Co., 
bleeder turbine, 191 
coupling, 169 
emergency governor, 142 
expansion joint, 183 
geared-turbine and generator, 165 

adjustment, 139 
impulse turbine, 48, 84 

TURBINE nozzles, 50 


ble-flow, 69 

single- and double-flow, 68 
single-flow, 67 
ON governor hunting, 128 

Westinghouse Electric & Mfg. Co., 
ON reduction gear alignment, 164 

thrust bearings, 99 
reaction turbine, cross-compound, 66 

single-double-flow, 65 

double-flow, 65 

steam conditions, in "Power," 280 
Windage losses, 17 
Wing, L. J. Mfg. Co.. 72 
Wheel, 25 
Work, perfect turbine, 15 


Corp., condenser graph, 181 
installation with barometric-jet con- 
denser, 178 



■^Y OlVISJOfi 





in USA 



3 9358 00012962 4 


prac tlce 

Croftf Terrell Williamst 1880- 
Steam- turbine principles and 

/ Terrell Croft, editor. — New York 

McGraw Hill, cl923« 

xi, 347 p. : ill. ; 21 cm. — 

(Library of power plant practice)