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STEAM-TURBINE
PRINCIPLES AND PRACTICE

TERRELL CROFT, Editor
CONTRIBUTORS

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

BOOKS ON PRACTICAL
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By Terrell Croft

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STEAM-TURBINE
PRINCIPLES AND PRACTICE

\J^

„lh^-^

^ TERRELL-CROFT, Editor

CONSULTING ENGINEER. DIRECTING EfTOINEER, TERRELL CROFT ENGINEERING CO.

MEMBER OP THE AMERICAN SOCIETY OF MECHANICAL ENGINEERS.

MEMBER OF AMERICAN INSTITUTE OF ELECTRICAL ENGINEERS.

MEMBER OF THE ILLUMINATING ENGINEERING SOCIETY.

MEMBER AMERICAN SOCIETY TESTING MATERIALS.

NORTHEASTERN U^VERSITV

Ss^y Division

First Edition
Fourth Impression

McGRAW-HILL BOOK COMPANY, Inc.
NEW YORK: 370 SEVENTH AVENUE

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

1923

C8S

PKINTED IN THE UNITED STATES OF AMEBICA

HE MAPLE PRESS - YORK PA

PAY CHV^^

PREFACE

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

vii

Vlll PREFACE

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.

PREFACE IX

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.

ACKNOWLEDGMENTS

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
text.

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.

CONTENTS

Paqb

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

STEAM-TUEBHE
PEINCIPLES km PRACTICE

Turbine
Wheel ■■

DIVISION 1

STEAM-TURBINE FUNDAMENTAL PRINCIPLES

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.

doiler

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

STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 1

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

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

under it several turbines were built.

Reducfhn Gears

r-~^

.Turbme Wheel

Pebfli;-

■Mortars

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]

FUNDAMENTAL PRINCIPLES

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.

mil-

of a

Shaft- . Vaned
Wheel'.

'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

Oav

n»\/ic!AH

4 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 1

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]

FUNDAMENTAL PRINCIPLES

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
waterwheel.

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

6 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 1

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.

Arms

I-Lawn Sprinkler

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

/////}/////'///^7/^.

I- No Jet

'Deflection,
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]

FUNDAMENTAL PRINCIPLES

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.

■5Lb.

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

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.

Supply

&z:z::::;::;;4~^i^

:zir:_-_ )

^' ^^^'^"^^

^^^

^^^^^

''Entering

■Plug.

Steam

'Balanced
Forces

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.

8 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 1

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

Sec. 10] FUNDAMENTAL PRINCIPLES 9

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-
pounds.

10 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 1

^§-A

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

!£; Lb. Per Horse power-Howr

|ik'i|ji,ii'u'JiiVil|ii'i'iliMi4i'ii|i|i|liii;ii'|i|^'|lMivX^

1.4 Ln+ropy 1.5

Constoint

Qudli+y Lines

Lines Of Constant
6ao|e_JPre55u

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

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

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

Sec. 10] FUNDAMENTAL PRINCIPLES 11

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 = 1010 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).

12 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div 1

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) '-^Biock^ 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]

FUNDAMENTAL PRINCIPLES

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
HrOr44TtPer5ec.

'mBm

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

^@&.^

Frame

yelocify Of

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

Rotating
Disk--

?^???^?;^^^^^^^^^^

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).

14 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 1

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.

Posfin
(yrouni.

■ 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.

wm^mm

^'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:

(11)

Kinetic energy of jet =

64.4

(ft.-lb.)

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

(ft.-lb.)

Sec. 13] FUNDAMENTAL PRINCIPLES 15

Hence,

(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

16 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 1

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]

FUNDAMENTAL PRINCIPLES

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.

18 STEAM-TURBINE PRINCIPLES AND PRACTICE [Dw. 1

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
jet.

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 /;« = 0
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

Sec. 17] FUNDAMENTAL PRINCIPLES 19

then Ve = 0 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

20 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 1

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 conversion 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.

Sec. 18] FUNDAMENTAL PRINCIPLES 21

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 =
0 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,

QUESTIONS ON DIVISION 1

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
only?

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.

22 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 1

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?

PROBLEMS ON DIVISION 1

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?

DIVISION 2

STEAM-TURBINE NOMENCLATURE AND
CLASSIFICATION

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.

Mouth-

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
service.

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

24 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 2

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
B/acfes
Fig. 24. — Convergent nozzles in
diaphragm.

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

Sec. 22] NOMENCLATURE AND CLASSIFICATION

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
wheels.

26 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 2

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

I-SideElevoi+ion

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

[<-... ./5/:..»| 'fPipe

Tap, Drain

From

Casing

<--8-

]!■ 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

Y-IO'^A
f" Pipe Open To
Atmosphere

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).

Sec. 25] NOMENCLATURE AND CLASSIFICATION 27

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.)

28 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 2

L/^

Sec. 28] NOMENCLATURE AND CLASSIFICATION

Spirvrl Mern'n^hone.
Gears - ■ ' '

Low-Speed Shaff.^
'Driven

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

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

Classifi-
cation

Class with
respect to

Class
No.

Class

Illus-
tration

Blading or nozzle ar-
rangement

2
3

Impulse.
Reaction.
Impulse-and-reaction.

Fig. 33
Fig. 35
Fig. 83

Service or steam con-
ditions

4
5
6

7
8

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

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

Single

10
11

Single pressure and velocity stage
(axial flow).

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

Fig. 41

Fig. 53
Fig. 42

Velocity stages
only

Single-pressure, several velocity
stages.

Fig. 33

Reactions tages
only

Many reaction stages.

Fig. 30

Pressure stages
only

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

17
18
19

Axial flow.
Tangential flow.
Radial flow.

Fig. 41
Fig. 42
P'ig. 43

Division of flow

20
21
22

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

Fig. 44
Fig. 45
Fig. 46

30 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 2

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

Carries

Coupling Casing -■■■'i^

Shaft Exfentton
Carries
..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

Sec. 31] NOMENCLATURE AND CLASSIFICATION

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

-yr^e/

Nozzle x->\

(CCCCCC

mSsM

nnro

•Stator
I- Circular Section

I- Longitudinal
Section

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

32 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 2

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
blades.

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

Sec. 32] NOMENCLATURE AND CLASSIFICATION

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
pressures.

Pressure.::-
Oages'
.-Steam

Fig. 36.

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

■: Fixed Blades.

Blades
I-Cylindrica!

' 'Blading Straightened
Out Into A Plane

' • • - -Plane Surface Of Section.

I- Transverse Section

E- Longitudinal

Section

Section
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
3
(

34 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 2

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
0 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.

Sec. 36] NOMENCLATURE AND CLASSIFICATION

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.)

Law-Pressure
Steam Inlet-'

Steam Divides Here-,

High-Pressure
Steam Inlet.

^-- -Exhaust To Condenser

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

High-Pressure
Sfeamlnlet'^x

-Exhaust
Outlet To
Condenser

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

36 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 2

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

Nozzles-:

"^^ In Three Sefs\Of

Nozzles 1

u •■■

^v,n^nozz,e^^^

^■■^

-•

—

^■

—

.— 1

-J

1 '

' '

' 1

II ,

velocity (jfaph

1 .1

1 1 1

1 1

[

1 1

■

Velocity G'ra

Velocity GVaph

IT 1 Ml

.-Velocity Drops

siL

^t"

Ix'

L..

•^

—

-, 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

Sec. 40] NOMENCLATURE AND CLASSIFICATION 37

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
stage.

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.

38 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 2

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

.'Steam-Supply
'<i Pipe

Blades-
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

Buckets.

•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

Sec. 44] NOMENCLATURE AND CLASSIFICATION

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
generator.

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

Moving
Blades.

; Live-Steam Inlet
\ ,■ Fixed Blades

.'Fixed Blades

Moving Blades^ Live -Steam
Inlet^

Blading'
^'^—E/.haust-Steam Outlet

Fig. 44. — Elementary single-flow reaction
turbine.

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

40 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 2

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

Outlets-
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.

QUESTIONS ON DIVISION 2

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
prevented?

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?

Sec. 46] NOMENCLATURE AND CLASSIFICATION 41

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
sketches.

DIVISION 3

STEAM-TURBINE TYPES AND CONSTRUCTION

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-
rangement

(AJableiS)

Staging Or

Cylinder

Arrangemen+

(C. Table V6)

Type

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

Re-Entry

Axial Flow
Tangential
Flow

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

Impulse

Veiocity-5taged

Curtis

Single Entry

Axial Flow

Pressure-Stageol

Roteaii

High-Pressure, Condensing Or

Velocity-And-
Pressure-Stageol

Curtis

Non-Condensing; And Low-
Pressure Condensing

Composite
Staged

Curtis And
Rateoiu

Single
Cylinder

Parsons

Single-Flow
Double-Flow
Single-And-Double Flow

High-Or Low-Pres'^ure,
Condensing

Reaction

Compounded

Parsons

Cross

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

High- Pressure Condensing

Tandem

Single
Cylinder

Curtis And
Parsons

Single Flow
Double Flow
Single-And Double Flow

Hiqh-Or LowPressure,
Condensing

Impuhe-And-
Reaction

Compounded

Curtis And
Parsons

Cross

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

High-Pressure Condensing

Tandem

Sir
FI

gle-And-Double

Sec.

TYPES AND CONSTRUCTION

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)

tloving
blades

Sfaiionaru
blades

Lxhaust,
Diaphragms.,

rOe Lava I
Type Or
jingle t>tage

IVelocity
Staged
Curtis Type
(One Pressure
Stage

Containing
Two Velocity
Stages)

ni-Pressurc-And-
Velocity Staged
Curtis Type
tTwo Pressure
Stages^Each
Containing
Two Velocity
Stages)

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 .

44 STEAM-TURBINE PRINCIPLES AND PRACTICE [Drv. 3

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)

f:r\

fiozzh
valve N

Nozzle,N
I

Fan Casing,

[Fan

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]

TYPES AND CONSTRUCTION

■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
sections.

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
TC.

46 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 3

Nozzle
Valves..
I

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.)

,'Ccrs/h^

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

Sec. 53] TYPES AND CONSTRUCTION 47

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

48 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 3

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

£x/?aust-

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

Moving
Blades^

'xhausf
Nozzle

Peversi'ngr^
Chamber

^- -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]

TYPES AND CONSTRUCTION

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

50 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 3

(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

Reversi'ngr
Chamber

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

Fig.

■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]

TYPES AND CONSTRUCTION

.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-'

Heer

Flangfe-''

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.

52 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 3

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.

fxhrcfsfCase

Inspection Hole.,

Stuffing Box.-.

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

Steam
Exhaust

Ball-
Beanng
Step
'Thrust
T-. Collar
Steam
Inlet
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]

TYPES AND CONSTRUCTION

.'•tnifht Pressure

Exhaust-
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

'Casing

Nozzle-'

■////■///A

Moving
Blades -

H-Longitudinal Section

Direction Of F
Rotation- -

ill

I-Cyllndrical Section

Sfaiionary
Reversing
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.

54 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 3

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. ''

.■Nozzle

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]

TYPES AND CONSTRUCTION

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
Nozzle

'Blades

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
View

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

56 STEA M-T (/RHINE I'lilNCIPLES A ND I'liA CTICE [ Div. 3

(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

Bearing

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

I'OrMore
Clearance,,

■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]

TYPES AND CONSTRUCTION

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

58 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 3

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

Governor-;

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

Diaphragm packing^' ^''°'"
rings

Steam
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]

TYPES AND CONSTRUCTION

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- ;

ill

Sil

^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).

60 STEAM-TURBINE PlllNCIPLES AND PRACTICE [Div. 3

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

■Governor

p.erM Byp.S5 Pfpes .'■■^fJSh^i:

Diaphragm

Shaffy

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]

TYPES AND CONSTRUCTION

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

hfmosphenc
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

"ooiinq
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

62 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 3

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
ForLiftingf

Exhausf-

End Cast/ngf - - -j^

Exhaust
Cavity — "

Carbon
Packi'nef-

(Diaphragm Cover

5team-Enc( Casting

'Impulse Bucket

''Oil Baffle
■Steam Supply Valve

'Live Steam Cavity
''Nozzle Valve
'Nozzle

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]

TYPES AND CONSTRUCTION

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
Gland),

By-pcrss Vafve
Baiance Piston^-

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

Co.)

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

64 STEAM-TURBINE PRINCIPLES AND PRACTICE [Drv. 3

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]

TYPES AND CONSTRUCTION

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.

Co.)

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^

Bearina

.Relief Va/re

'£i</?aus/

Exhaust

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

66 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 3

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-
Cylinder

■ Hlgh-PressuKe
Cylinder

(Governor:

■Lxhausf To
Condenser-

iiJ \:5feam
■\X;;5ypp!y

■Founcfafi'on

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^

Low-pressure
Cylinder

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.
Co.)

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]

TYPES AND CONSTRUCTION

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

68 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 3

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

hmliary Steam Infef

M'efValve^-:

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.

Safofy
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).

Sec. 69] TYPES AND CONSTRUCTION 69

QUESTIONS ON DIVISION S

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
adapted?

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
turbines?

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
turbine.

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
turbine?

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

70 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 3
70. Table by Manufacturers Showing Steam Conditions,

Manufacturer

Fig.
No.

Steam conditions

Class or description

Alberger Pump & Condenser Co.

High-pressure, non-
condensing

Curtis, impulse, sin-
gle flow

Allis-Chalmers Mfg. Co

All

Parsons, reaction, sin-
gle-flow

Calling Turbo-Blower Co

High-pressure, non-
condensing

Impulse, single-flow

Coppus Engineering & Equip-
ment Co.

High-pressure, non-
condensing

Impulse, single-flow

High-pressure, con-
densing and non-
condensing

Impulse, single-flow

De Laval Steam Turbine Co

High-pressure, con-
densing and non-
condensing

Impulse, single-flow

All

Impulse, single-flow

High-pressure, non-
condensing

Curtis, impulse, sin-
gle-flow

All

Curtis, impulse, sin-
gle-flow

High- and mixed-
pressure, condens-
ing

Impulse, single-flow.

and double-flow

All

High-pressure, non-
condensing.

Impulse, re-entry

Moore Steam Turbine Corp

High-pressure, con-
densing and non-
condensing

Impulse, single-flow

Parsons Marine Steam Turbine
Co.

High- and low-pres-
sure

Parsons, reversing

* Oil relay governors used on some large machines.

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

Sec. 70] TYPES AND CONSTRUCTION 71

Classes and Approximate Ratings of Steam Turbines

Approxi-
mate
ratings

Type of
staging

Governor

Glands

Notes

10-50 hp.

Velocity

Pickering

Lantern

Centrifugal-pump drive

300 kw.
and up

Reaction

Throttling and by-
pass

Water packed

Turbo-generator, direct
drive

1-25 hp.

Velocity

Direct throttling

Metal packing

Blower drive

2-50 hp.

Single

Pickering

Stuffing box

Blower and pump
drive

1-600 hp.

Single

Direct throttling

Metal packing

1-600 hp.

Velocity

Direct throttling

Metal packing

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

50-15,000
hp.

Pressure

Direct throttling*
or oil relay

Carbon packed

10-400
hp.

Pressure-and
velocity

Direct throttling

Steam packed

Mechanical drive

100 kw.
and up

Composite

Direct, t steam, oil
relay

Steam laby-
rinth

Director gear-con-
nected for generator
drive

300-4,000
hp.

Composite

Oil relay

Steam laby-
rinth

Turbo-compressors

Up to
4,500 hp.

Velocity

Direct throttling*
and oil relay

Carbon packed

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

Composite

1-800 hp.

Single

Direct throttling

Carbon packed

Direct connected for
pump, generator, and
blower drive

Up to
4,500 hp.

Velocity

Direct throttling
and oil* relay

Carbon packed

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

Composite

Reaction

Marine service

72 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 3
70. Table by Manufacturers Showing Steam Conditions, Clas-

Manufacture

Fig.
No.

Steam conditions

Class or
description

Power Turbo Blower Co

High-pressure,
condensing

non-

Impulse, single-flow

Ridgeway Dynamo & Engine Co.

High- and mixed-
pressure

Impulse, single-flow

High-pressure,
condensing

non-

Impulse, single-flow

B. F. Sturtevant Co

High-pressure,
condensing

non-

Impulse, tangential-
flow

Terry Steam Turbine Co

High-pressure,
condensing

non-

Impulse, tangential-
flow

High-pressure,
densing

con-

Impulse, single-flow

Westinghouse Electric & Mfg.
Co.

High-pressure,
condensing

non-

Impulse, re-entry

80.

82, 83,
84,85

All

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

L. J. Wing Mfg. Co

High-pressure,
condensing

non-

Impulse

Oil relay governors used on some large machines.

Sec. 70] TYPES AND CONSTRUCTION 73

ses, and Approximate Ratings of Steam Turbines — Continued

Approxi-
mate
ratings

Type of
staging

Governor

Glands

Notes

1-20 hp.

Single

None

Metal packing

Blower drive

500-5,000
kw.

Composite

Pneumatic oil re-
lay

Carbon pack-
ing

Generator drive

Up to
300 hp.

Velocity

Direct throttling

Steam laby-
rinth

Direct-connected pump
drive

5-350 hp.

Single

Direct throttling

Stuffing box

Direct pump and gen-
erator drive

15-500
kw.

Single

Direct throttling

Carbon packed

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

1-1,500
hp.

Composite

Direct throttling
or oil relay *

Carbon packed

2-640 hp.

Single

Direct throttling

Water seal

Mechanical drive

500 kw.
and up

Reaction, im-
pulse-and-
reaction

Steam or oil re-
lay, and inter-
mittent

Centrifugal
water-packed

Turbo-generators

1-20 hp.

Single

None

Stuffing box

Blower drive

DIVISIOX 4
STEAM-TURBINE INSTALLATION

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]

STEAM-TURBINE INSTALLATION

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- - ' -

Atmospheric

Exhaust Reiki

Yalye--

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.^ \^^[\
ForAc/Justiny
Turbine •■_ Pressi/re. h

'Extiausf

SyPcns-

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

76 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 4

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]

STEAM-TURBINE INSTALLATION

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

bedplate

\-''0.

- 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

78 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 4

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
Erection.

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

Sec. 79] STEAM-TURBINE INSTALLATION 79

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
problem.

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.

80 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 4

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

Sec. 81] STEAM-TURBINE INSTALLATION 81

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
place.

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.

82 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 4

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

Sec. 85] STEAM-TURBINE INSTALLATION 83

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

84 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 4

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^.

.'Center
[Support

Cast-iron
Casing \

■Bearing

Bedplate

•Shaft

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
alignment.

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.

Sec. 87] STEAM-TURBINE INSTALLATION ' 85

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.

QUESTIONS ON DIVISION 4

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?

DIVISION 5

STEAM-TURBINE SHAFTS, BEARINGS, AND PACKING

GLANDS

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

couplings.

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]

SHAFTS, BEARINGS, AND GLANDS

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

^m^

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

Col la r--'

^nd / 1^ ^a, .1. . '"Mua^ ^

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.
93.

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:

5a'arr^

D,^rnn

"^^^®

^\ Jet A

Thrust
Co.'^ar

^"^

^rz

S'^

Spina ir

■X," '^

mij^^^^^^i

^^V "^'''"'9

Geary

Of/ Sling'

; '-Labyrir.

C'!and^^^//lll^

\' ■

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

88 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 5

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
sections.

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.)

Steam-turbine
bearings

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]

SHAFTS, BEARINGS, AND GLANDS

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

Spherical
5eaf

Collar
Bolt

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.

90 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 5

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
Lining

babbitt

Oil

Groove

Lower
•Lining

<1>

-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,

Sec. 92] SHAFTS, BEARINGS, AND GLANDS

<-F

Casf-Iron ^

Lining "^;,

l-Tubes Assembled

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

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.]

Ot/
Vapor

AXIS Of
Shaft Ands^
Bearing
before
Bending

QitVenf
Bearing Cover

Pedesfaf

'mw/wA

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

92 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 5

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-
tion.)

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]

SHAFTS, BEARINGS, AND GLANDS

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

'S/?crf/-

.'Spherical Seat

-Bearlnof Coyer

Upper Half Lining

I-Partial Longi+udinal
Sec+ion

doffpm / H-Sec+iion X-X

Half Linincf jv,.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.

94 STEAM-TURBINE PRINCIPLES AND PRACTICE [Drv. 5

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]

SHAFTS, BEARINGS, AND GLANDS

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

Shims

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

96 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 5

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
Moving
Thrust Cage
And Adjusting
Position Of

Rotating

Element

Thrust Cage .'

Moved Axial ly '

By Adjusting Bolt, A,

And Prevented From

Turning By Key, K.

Roller
Bearings

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]

SHAFTS, BEARINGS, AND GLANDS

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.

Co.)

the side opposite from the direction of thrust to restrain any endwise
movement of the shaft. The bearing blocks, F, (as shown in ///), rest
7

98 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 5

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

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]

SHAFTS, BEARINGS, AND GLANDS

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

■->A

^^^-^^-c.K'^iiJnr^

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-

100 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 5

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.

Sec. 102] SHAFTS, BEARINGS, AND GLANDS 101

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.

102 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 5

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
below.

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]

SHAFTS, BEARINGS, AND GLANDS

103

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
Packing

Stuffing
' Locknuf Box

Wafer Def/ector
■ EKhausf Felt Washer'

Case

Bearing .
Case-'

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

104 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 5

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

\<-'"-^"i'Toi'

B<' , E-5ection IH-Skivcd

I-Elevoition A-B Or Lap

Joint

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]

SHAFTS, BEARINGS, AND GLANDS

105

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

balance-Pisfon

Clearance Clearance f^mofS '^"/3^

Pig.

Siaflonary
Casing

"•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.

106 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 5

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

Turbine
{Runner

■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]

SHAFTS, BEARINGS, AND GLANDS

107

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.

Diaphragm

Packing
Plate - •

,Garter
' 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.

Packing
Chamber

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

108 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 5

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

A<-

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

Sec. 114] SHAFTS, BEARINGS, AND GLANDS

109

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
6lanc(~Wafer
Pressure Gaofe'

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

'--Drain To
Zxhausf

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

110 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 5

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.

Sec. 118] SHAFTS, BEARINGS, AND GLANDS

111

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

...r/oaf
Valine

.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-

112 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 5

'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.)

Sec. 119] SHAFTS, BEARINGS, AND GLANDS

113

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-

S/7afA

Spiral Spring
(Garter Spring)

ITwo Chambers Each Con+alning|
Two Packlnoj Rings

^ . , Gland Chamber B
Axial '■ P7:

Clearance--

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

Carbon '*•

Piece

Jiingf No. 2

K-Two Chambers Each Containing
One Packinoj Rinqj

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

Dra/nOrLeal<-Off

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
turbine.

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

114 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 5

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
sections.

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

Sec. 112] SHAFTS, BEARINGS, AND GLANDS 115

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.

116 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 5

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

Perfect

fit between

Carbon And

Wooden Holder-''
Wooden'
Model

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

Board

Fig. 124. — Hacksaw used to decrease
the diameter of a carbon packing ring.
(E. H. Thompson in Power, Sept. 21,
1920.)

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,

Sec. 123] SHAFTS, BEARINGS, AND GLANDS 117

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.

QUESTIONS ON DIVISION 6

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
turbines?

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
bearings.

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
determined?

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?

118 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 5

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
vaporizing?

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.

DIVISION 6

STEAM-TURBINE GOVERNORS AND VALVES

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

Governor
Vctlve --■

"Spent" rCca^vf;W

Shaft

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

119

120 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 6

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--
G

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

riy-Ba/l:
Governor

Nozzle--

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]

GOVERNORS AND VALVES

121

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 \ .'KnifeEdge

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

122 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 6

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

Governor
Lever -

'Governor
Butterfly Valve
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
valves.

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]

GOVERNORS AND VALVES

123

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
sections.)

Classi-
fication

Classified with
respect to

Class
No.

Class or
description

Illus-
tration

Actuating force.

Centrifugal.

Fig. 136

Centrifugal and inertia.

Fig. 156

Air pressure.

Fig. 157

Method of valve

Direct.

Fig. 130

control.

Mechanical indirect.
Indirect or relay:

Fig. 159

(a) Steam relay.

Fig. 143

(6) Oil relay.

Fig. 142

Valve arrangement

Throttling.

Fig. 130

or steam control.

Varying nozzle area
(multiple valve).

Fig. 152

Intermittent.

Fig. 147

Bypass.

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.

124 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 6

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,

Ik- V Centrifugal- Weight Arms

ernorSpindi
'jll-BeanngSa
: Bel /-Crank lever J)

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

Lock Nut ^
Thrust Bearing^
Speed-
Adjusting NutW

•ustlng Nut\,
■External Spring P

y^-\/alve r 0

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
friction.

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

Sec. 131] GOVERNORS AND VALVES 125

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

126 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 6

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]

GOVERNORS AND VALVES

127

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

''Strainer
'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.

128 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 6

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.

Fly-Bails-

>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]

GOVERNORS AND VALVES

129

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

Bell
Crank

Fig. 136. — Emergency- and main-governor-mechanism assembly of Type-6 Sturtevant

turbine.

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

130 STEAM-TURBINE PRINCIPLES AND PRACTICE [Drv. 6

distance which the piston extends when it flies out. The
clearance between the tripper, B, and the rotating element

r Adjusting Plug
.'Spring
^^-D/rect/pn
^^^ff^otation

:" ji^^iearance
( 'Tisfon^A
''Stop Bushing

Turbine
Shaft

Tripper

^Valve
Spring
''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

■Trigger
''TripLever
' Trip -Lever Siiaft
I-Section A-A

;' ''Governor
Cup

'^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
governor.

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]

GOVERNORS AND VALVES

131

adjusted to trip at about 10 per cent, above the normal
running speed. The emergency governor should be tested

leaf
Spring-

Depressions To Hold Spring

\ Slot; .-Motion Limiting Stud

Finger-^

Plvof
Stud

Emergency- Vaive
Operating S/?a ft'.

Emergencif-
Vaive Operating
Finger

Governor ^
Vise -''

Fig. 139. — Ring-type emergency
governor used on the smaller Terry
turbines.

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

132 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 6

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

'Synchron/i.
Oil Pump-

's. '•■0/IUncfer
:■ 3£To40Lb.
PerSq.In.
I Pressure
'Sieam-End
Bearing Case

Steam
'"eovernor- Supply
Steam Valve
^'' Steam Chest

'Spriny

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]

GOVERNORS AND VALVES

133

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.

134 STEAM-TURBINE PRINCIPLES AND PRACTICE [Bw. 6

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

2570

V-2Z40^

270

-Time

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]

GOVERNORS AND VALVES

135

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
devices.

Oil Outlets
To Governor-
Operating
: Piston "'^

136 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 6

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

ports.

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.
145.

138. The Westinghouse Type Of Centrifugal Oil-relay
Intermittent Governor is shown in Fig. 147 and the valves

Sec. 138] GOVERNORS AND VALVES 137

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, 0 (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, 0 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
positions.

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

138 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 6

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;
5fop/alye-~

Coyer-
Main Goyernor Spring.

Coyernor Ball- - -V

Piyof Of nockingl
Leyer On Ooyernor •;
Leyer- .

* ^Operating
Phfon

^ynchronmng
MotQf

Hanctwheel-

Speecf Changer
Or Synchronizing.
Spring-.

Worm;
Wheel

Worm-'
^-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

Sec. 139] GOVERNORS AND VALVES 139

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
valves.

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

140 STEAM-TURBINE PRINCIPLES AND PRACTICE [Drv. 6

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

strainer

Secondary
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]

GOVERNORS AND VALVES

141

Fig. 149. — Enlarged view of the Westinghouse relay valve of Fig. 147.

142 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 6

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

'•EmergencLf
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]

GOVERNORS AND VALVES

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 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 6

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

7////////////////////////////////////^^^^

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]

GOVERNORS AND VALVES

145

■Cam Shaft

chiefly by varying the number of nozzles to which steam is
admitted.

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,

Section
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.

146 STEAM-TURBINE PRINCIPLES AND PRACTICE [Drv. 6

Auxiliary Of
Synch ron iz in^
^^ Sprlncf^

Limit
Switches

Motor

Contra!

Wires^

tiynzhronizincf Motor;'

Fig. 154. — Vertical centrifugal governor used on large-capacity General Electric Co.

Curtis turbines.

Operatincf

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.

Sec. 147] GOVERNORS AND VALVES 147

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

148 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 6

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.

Parallel

Parallel
Link

'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]

GOVERNORS AND VALVES

149

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

150 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 6

-—Crossheoict

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
vacuum.

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]

GOVERNORS AND VALVES

151

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

152 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 6

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
kerosene.

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

Sec. 152] GOVERNORS AND VALVES 153

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,

154 STEAM-TURBINE PRINCIPLES AND PRACTICE [Drv. 6

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]

GOVERNORS AND VALVES

155

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
needed.

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.)

156 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 6

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.

Bonnet

{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
Accessories.

Sprfngr,

Pipe thread
Connection,

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

Sec. 155] GOVERNORS AND VALVES 157

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-

158 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 9

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

5hee-f-,Mefa/

Cylinder.^

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.

QUESTIONS ON DIVISION 6

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
governor.

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
function.

11. At about how much greater than normal speed is the emergency governor usually
Bet?

Sec. 156] GOVERNORS AND VALVES 159

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-
advantages?

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-
ages?

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
governors?

30. What is the general construction of most steam strainers for turbines? What is
their function?

DIVISION 7

STEAM-TURBINE REDUCTION GEARS AND
COUPLINGS

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

160

Sec. 158] REDUCTION GEARS AND COUPLINGS

161

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.
11

162 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 7

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

•5|

fl o

^ S-

1
<

■1-

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.

Sec. 159] REDUCTION GEARS AND COUPLINGS

163

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

Gear

.'Supply Line To Of I Cooler Case Cap-.
\ OearBearing, /^

I Gear thrive / Q'f

' T fhmp

•Supply Line ToBearings
From Oil Cooler
Pinion
/'Bearing'

Supply Line
• To Oil Spray
\ Tube
^ Oil- Pump Coupling

Oif
Slinger

~- Overflow
From Bearings

Inner
OilRing

^- - 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
overflowing.

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

164 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 7

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:

Sec. 161] REDUCTION GEARS AND COUPLINGS

165

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 ^^ 0<Co//ars ^ (Cd

~«, Pinion-^ \

Gearwheel

-Turbine
I Bearing ^

Turbine-Wheel •
Casing

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.

166 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 7

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\
Speeci

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

Sec. 163] REDUCTION GEARS AND COUPLINGS

167

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,

Pinion
Shaft •;

■Oil outlet
to Bearings^

Valves

Oil-"

Strainer

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

168 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 7

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 '

'Ri'gr/of
Qqupling

Outboard- beanncf
Journal

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

Driven
Coupling

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.

Sec. 165] REDUCTION GEARS AND COUPLINGS

169

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.
\ Bushing ^ , -■■

. 'VrWing Pins-^

E-^ End Yiew_

Wire".

dross 'dushing ' ^J

1-Longitudinal Section

Section
tK-K

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- - . .^
f/eeyeS'^/^

;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.

170 STEAM-TURBINE PRINCIPLE.'^ AND PRACTICE [Dj\\ 7

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
parts.

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

Sec. 167] REDUCTION GEARS AND COUPLINGS

171

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

Down

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

STANDARD TABLE OF DIMENSIONS TO BE OBTAINED EACH
TIME ALIGNMENT BETWEEN TWO SHAFTS IS CHECKED

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

Position

Inchas

Position

Inches

0.124

o.ooa

R.H. Side

0.124

R.H. Side

0.005

Down

0.124

Down

0.008

L.H. 5ide

0.124

L.H. Side

0.008

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.

172 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 7

QUESTIONS ON DIVISION 7

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
against.

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.

DIVISION 8

STEAM-TURBINE REGENERATORS AND CONDENSERS

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

High-Pressure

5feam
From Boifer

Fig.

Low-Pressure

5feam From

engine

4-SidV Elcvrt+ion Showing Piplnoj Arrangement

179. — A typical Rateau regenerator or accumulator for use with low-pressure
turbines.

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
jacket.

173

174 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 8

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, 0 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.

Sec. 169] REGENERATORS AND CONDENSERS

175

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

■boilers.

High- Pressure
Header,

High-Pressure 5feam For
Auxiliary Supply To Turbine-.

Generator

Fig.

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-

176 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 8

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.

Sec. 172] REGENERATORS AND CONDENSERS 177

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
12

178 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 8

wwMi^4w^ ':

Fig, 181. — Arrangement of equipment in a turbine-driven plant showing surface con-
denser and auxiliaries.

Baromefric

Overflow' vi] . ^ uuu..^ i p f f n

Cold Y^aferSucIfoiv^ / Entminer Circulafinq Pump Varui'im Pum,
Hof Water 5ucf ion ^ vacuum ^um,

Pump

Fig. 182. — Steam-turbine installation with barometric-jet condenser, C, and cooling
tower, T. (Worthington Pump and Machinery Corp.)

Sec. 173] REGENERATORS AND CONDENSERS

179

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->^^-:)

Mulfi-Jet
Condenser- -

-Wafer Discharge From
Pump To Condenser

'Centrifugal Circulating-
Water Pump

n>,M.>w^w^m^w~-^w.~WA^>>

■ '';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
required.)

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-

180 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 8

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.

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-

Sec. 174] REGENERATORS AND CONDENSERS

181

Square Feet Surface Per 1000 Lb. 5 + cam
100 F 100 500 400 500

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1 ftO* 110^ 100* 90^^ 60-
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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.)

182 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 8

Lever for

Opening

Valve By

Hand

Supporting
Webs

Sfeam
Inlet-.

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
Chamber

Thermometer
Connection

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-

Sec. 174] REGENERATORS AND CONDENSERS

K 6^~ M

r/angre

.Wafer
/Inlet

183

I-Vertica\ Section Showing
General Assembly

Conofensif
rianffe

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
chambers.

184 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 8

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.

QUESTIONS ON DIVISION 8

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?

Sec. 174] REGENERATORS AND CONDENSERS 185

PROBLEMS ON DIVISION 8

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
steam.

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?

DIVISION 9

HIGH-PRESSURE, BLEEDER, MIXED -PRESSURE AND
EXHAUST-STEAM TURBINES

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; ' ]

E'5

A) -1

f V^

|:;l.ovy

-Pres-jilre

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
engine.

Note. — When Both Heat And Power Are Supplied By The Same
Power Plant, it is economical to generate the steam (which will be

186

Sec. 176]

SPECIAL-SERVICE TURBINES

187

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.

Case

Exhaust steam avail-
able

Exhaust steam
needed

Turbine used

None

None •

Condensing turbine

Always more than
turbine needs

None

Low-pressure turbine

Sometimes more —
sometimes less than
turbine needs

None

Mixed-pressure tur-
bine or low-pressure
turbine and regen-
erator

None

Always more than
turbine will supply

Non-condensing tur-
bine and reducing
valve

None

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

188 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 9

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
turbines.

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]

SPECIAL-SERVICE TURBINES

189

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
Steam 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.

190 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 9

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

Live-Sfeam
Heaafer-

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

SPECIAL-SERVICE TURBINES

191

operate: (1) Condensing, (2) non-condensing, or (3) partly
condensing and partly non-condensing at the same time. Under

bleeder Valve

' Oovernon

High-
Pressure i.\y
Steam
Inlet

Exhaust To Condenser
Througt) Base

Generator

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

turbine.

some conditions it will operate almost wholly as a condensing
unit; under others, almost wholly as a non-condensing unit.

192 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 9

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- - '■•

Separator-

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]

SPECIAL-SERVICE TURBINES

193

194 STEAM-TURBINE PRINCIPLES AND PRACTICE [Drv. 9

Sec. 182]

SPECIAL-SERVICE TURBINES

195

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

196 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 9

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]

SPECIAL-SERVICE TURBINES

197

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
Turbine_

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

198 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 9

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

High-Pressure
Valve

Low-Pressure
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.-

Aufomafic
Travel

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]

SPECIAL-SERVICE TURBINES

199

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

Adjusting-
.■■Block

' 5econo/arL/
Valve-5fem
\ Link.

Opera fin^
Piston Rod..

To
Turbine

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

200 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 9

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

Indicator

Hand-Wheel For Lowering Or
Ro/i's/ngr Pressure Plate

Piston Chambers
"null ':

Peciprocatin^
Ln^ine

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]

SPECIAL-SERVICE TURBINES

201

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

,^Weights-

Lever-

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)

202 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 9

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]

SPECIAL-SERVICE TURBINES

203

■ ■ • ■ • ■ . ' . ■ ' 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'

^UUUULIUUUU

::]nnnnnnnnc:,
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.

204 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 9

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]

SPECIAL-SERVICE TURBINES

205

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.

206 STEAM-TURBINE PRINCIPLES AND PRACTICE [Dtv. 9

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]

SPECIAL-SERVICE TURBINES

207

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

I20

h^'-

ncf

— .

Si
c

eedin^

(0,000

Lb. Per

Hour,

/■

,.<

^ii>

(

) 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,

208 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 9

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
turbines.

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] =

Sec. 191] SPECIAL-SERVICE TURBINES 209

[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 0 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.

210 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 9

3.5

3.0

2.5

!.5

1.0

<?/

.7i?7

-.i

i./V

(

o<^

<i7

-f

S'^^

\ ^

^ /

^^^cS^

^.*.

'Pr

{My

/i/

'/>

-^z-

r>i

'.4/

....

...

0i5

1.0

Fioo

1.5

2^.

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 0 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]

SPECIAL-SERVICE TURBINES

211

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

Thrusf
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,

212 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 9

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]

SPECIAL-SERVICE TURBINES

213

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- -
Alfernafing
Current Bus

PSI

Non-Condensing
Rec iproca fing
Engine,

h=^

Alfernafirig Current
Generator

'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

214 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 9

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
loads.

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.

Sec. 194] SPECIAL-SERVICE TURBINES
Non-Conofensi'ngr Engine .. _

215

Flywheel- - . ^

Line 5 haft- ^

Motor-.

Belt-.

Three-Phase

Alfernafing-

Current L'me5\

f\\

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.

Operating
' Pisfon

Auxiliary

High-
Pressure
Steam
Valve

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.

216 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 9

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

l^mnoi

'■^Separating Receiver
^"Non-Condensing Engine

Flow
/a/ye

Condenser-'

■--Belf

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.

QUESTIONS ON DIVISION 9

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?

Sec. 194] SPECIAL-SERVICE TURBINES 217

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
constant?

8. A mixed-pressure turbine combines the functions of what two other kinds of
turbines?

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
governor?

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?

PROBLEMS ON DIVISION 9

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?

DIVISION 10

STEAM-TURBINE LUBRICATION

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

218

Sec. 196]

STEAM-TURBINE LUBRICATION

219

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-^

Wafer

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

220 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 10

.--Auxiliary Urn From Turbo Pump (Starting)

'^ ^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.

Crease

Grease^

{///('// ////{-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.)

Sec. 197] STEAM-TURBINE LUBRICATION 221

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

222 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 10

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.)

Cover--

Noxxle^
Casing)

-Casing
•Spindle Bushing
Spindle Collar
..-Impeller
Step Bearing

•aring Discs

Strainer
Submerged
In Oil Tank

-Turbine-driven auxiliary
(Allis-Chalmers Mfg. Co.)

lialn-Turbine
Steam Chest

Hoinci- Operated
bypass Valve

Connecfed To

Oil Pressure

5 y stem

Steam To

Auxiliary

Oil-Pump

Turbine

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]

STEAM-TURBINE LUBRICATION

223

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

224 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 10

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
possible.

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]

STEAM-TURBINE LUBRICATION

225

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.

226 STEAM-TURBINE PRINCIPLES AND PRACTICE (Div. 10

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)

^100

■^

1 n

■^

'vl'

^90

^•:

^'NJ

>^.

■^

^s,

Bai

C/7''

s^>

1 1 ',

_ . 'Styee t^n/nn .

s.<

v..

7'^

Treatment'

-''

**.

'«•,

s^Uj^..

--"^

B*N

s..

•-.,

eng

rime

,ln

Mon

+h5,

l5l

^ Tu

rblr

»e

)

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.''

Sec. 199] STEAM-TURBINE LUBRICATION

227

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
Discharge
Delivering
Filtered
oil To
TUrblne,
Reservoir

Duplex
Steam

Oil Pumpy

Oil

Pump
Suction
From
Fllten

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.

228 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 10

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 .

^l^---DirfL/-Oi/

In let

Pure-Oil
Compar-f-menf-'^

Helicai-Gear Dri've \:

Note-

The Spouts May
Readily Be Turned
To The Mo5i
Convenient
.Position

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

Sec. 201] STEAM-TURBINE LUBRICATION 229

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

230 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 10

(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
Company.)

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

Sec. 204] STEAM-TURBINE LUBRICATION 231

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

232 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 10

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]

STEAM-TURBINE LUBRICATION

233

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
maintained.

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.

Cylinder.

Inlet-
Va/ve 5tem
Oil Cup

> ^

-,^

(

-■■>

7 ~ *^

fflt-

-Relay
Piston

■GlancT

Inlet Valve Stem

Throttle Handwheef
Shaft Bevel Gear,

Throttle
Valve

Throttle
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.

This

234 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 10

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.

QUESTIONS ON DIVISION 10

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
differ?

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
oils?

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.

DIVISION 11
STEAM-TURBINE OPERATION AND MAINTENANCE

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.

235

236 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 11

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

Sec. 210] OPERATION AND MAINTENANCE 237

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
followed.

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-

238 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 11

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
Symptoms.

(Adapted from E. V. Amy, in Electric World, vol. 74, p. 1004)

Cause

How identified

Probable reason

What to do

Unbalance

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
unbalance.

(a) Sprung shaft.
(6) Improperly placed balance
weights.

weights. _

(d) Sediment in blades or buck-

ets.

(e) Corroded blades.

(/) Unequal heating of rotor

parts.
{g) Unbalanced forces due to

heavy distortional stresses.
{h) Shifting of conductors on

generator.
(z) Unequal generator air gaps.

As soon as vibra-
tion becomes ab-
normal, shut
down and investi-
gate. Remove
cause; rebalance
if necessary.

Poor
alignment

Vibration of variable
periodicity; slight
at no load, becom-
ing worse as load is
applied.

(a) Eccentric coupling.

(6) Unequal settHng of founda-
tion.

Check up and cor-
rect.

Bad
foundation

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-
plate.

Make foundation
solid; grout with
lead if necessary.

Loose
parts

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.

{d) Loose conductors on gener-
ator.

(e) Loose coupling or bolts.

Carefully examine
all bearings and
fastenings. Make
necessary repairs.

Internal
rubbing

Abnormal vibration
somewhat localized;
noise varying with
speed of machine.

(a) Revolving buckets coming
in contact with stationary
buckets.

ib) Insufficient casing clear-
ance.

{d) Thrust-bearing troubles.

Make repairs or
adjustments im-
mediately; may
result in serious
damage.

Sec. 213] OPERATION AND MAINTENANCE

239

Cause

How identified

Probable reason

What to do

Steam
troubles

Unusual noise near
the intake; failure
of the steam strain-
er.

(a) Water coming over with the

steam.
(6) Sediment in the steam.

irregular steani admission.

(d) Accidental closing of emer-

gency steam valve shifting
generator's load to other
machines.

Make necessary re-
pairs; test steam
for sediment,
acid, or salt.

Packing
troubles

Local vibration;
noise; heating of
shaft or packing
casing.

(a) Improper adjustment of

labyrinth packing.
(6) Packing rings too small for

shaft.

Make adjustment
or replace old
packing.

Oil
troubles

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.

emulsifying).

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

240 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 11

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

^-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
bearings.

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
ones).

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-

^Gland
Drain
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.

and

Sec. 215] OPERATION AND MAINTENANCE 241

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
prevented.

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
motion.

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.

242 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 11

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.

'//////////////////////////////////A
^^Concfensafe Pump

-Typical arrangement of piping, valves and auxiliaries for a condensing
turbine.

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
mercury.

12. Shut down the auxiliary oil pump and see that the main pump
keeps up the pressure.

Sec. 216] OPERATION AND MAINTENANCE

243

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. |

Time

STEAM PRESSURES

OIL PRESS.

TP%SE

a5vbb

Water Temp.

Supply

Si.g.-1

Stag*.2

Stegea

Belay

B«rir^

Inlet

Outlet

Inlet

Outlet

Vacuum

BoromlRemarKfl

izyo

1/3

■io

3./.t

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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.

244 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 11

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

Air

Pump Discharge., _^ Rece!.k?"^l°^^^^^%

—t ■;■,.■.■■ \ uuuru

k i^///////////////////////^

■ -Draff
Gage-.

■^age \ f/

■Sharp-Edged Opening

Board

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

Sec. 218] OPERATION AND MAINTENANCE 245

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
again.

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

246 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 11

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
resistance.

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
switch.

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).

Sec. 221] OPERATION AND MAINTENANCE 247

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

248 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 11

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.

Sec. 225] OPERATION AND MAINTENANCE 249

(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
covered.

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.

QUESTIONS ON DIVISION 11

1. State the three fundamentals of steam-turbine operation in the order of their
importance.

2. What precautions should be observed if successful operation of a turbine is to be
attained?

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
sketch.

10. Should the turbine rotor be turning when the steam is turned on to warm it?
Why?

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?

250 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 11

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?

DIVISION 12
STEAM-TURBINE TESTING

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
operates.

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

251

252 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 12

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]

STEAM-TURBINE TESTING

253

.■■Live Steam From Boiler

,;W

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.

-Thermomefer

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.

254 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 12

■Sfeam Fbw Mefer- ->

Thermomefer.

Calonmefen

^w^^^^^^?^^^^^^^^^^^!^m^^^^?5^^^^^^^^^:^^^

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.

Sec. 231] STEAM-TURBINE TESTING 255

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 thermometer and^tbgi
temperature of the steam as obtained from tjie steam 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

256 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 12

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
test.
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.

Sec. 234] STEAM-TURBINE TESTING 257

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

(pounds)

Hours duration of test
This method of determining the steam consumption is only

258 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 12

practicable where the turbine is operated (Fig. 237) in con-
junction with a surface condenser. The arrangement of tanks,

'Scale Platforms-

--; ^.r^rVerfical Sump

-A

fmmM/M/MM^Mkm

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

Sec. 237] STEAM-TURBINE TESTING 259

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
temperature.

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

260 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 12

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

Thro-tfle

Open Feed' loafer
Heater- ,

, Back- Pressure Turbine-
■Drlyen Boiler-Feed Pump

■Plat Form Scales-^ H-->;nng

V^/^j/// ^^^^ '''^ V'^^''' '^'/ ^^^^/^/7////

Circulating J
Pump --'

^Condensate
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

Sec. 238] STEAM-TURBINE TESTING 261

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

262 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 12

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

Inofi'cafoP

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 percentag^,^ 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]

STEAM-TURBINE TESTING

263

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-

18,000

16,000

HfiOO

3 12,000

10,000

8,000

Q. 6,000

3
«P

o
o

E
o

4,000

_ 2,000
o

\t^

Xfuafferi

^diffLohbL

ZOO
Load,

Tuirioaol

Zl i^

.91

16a:
15^

400 600

In Kilowatts

600

1,000

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

264 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 12

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 :

(35)

y^K =

t^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

TIME

LIVE STEAM

RJ>M.

liil»

In. Gage

Superheat,

•K Or .

<lirtlHyIa%

ST|^M

Genenitor

Kw.

ExcHer

Kw.

Net Power

Kw.

X:oe

t S3

,02,

txs

10,100

300

%900

1,8 00

4 3(., too

l-.jo

/S2.

/OS

73./

f.150

300

%6S0

1,8 00

3: 00

/fZ

fZ.i

10,100

300

%800

1,8 00

1:30

/4»

/Ol

f$.0

/0,IO0

3O0

1,80 0

Uoo

4:oe

/4f

91.B

10, ISO

300

9, 8 so

1.800

4:30

1 SO

n.e

10,100

300

1,9O0

1, 800

s:oo

/sz

100

tl.8

10,100

3 00

9,900

1,800

Average

1 S0.8 1 100.4 1 t2,.8 1 iO,lt8

300

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
32°F.)

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
kw.-hr.

(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.)

(36)

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.

Sec. 240] STEAM-TURBINE TESTING 265

(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 (ft..lb.perB.t.u.)

(43) Er = -^ (decimal)

266 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 12

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

Sec. 242] STEAM-TURBINE TESTING 267

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.

268 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 12

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.

(decimal)

total heat input per lb.
or, using symbols;

XT XT

(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 given
engine converts into mechanical work.\ It can be determined
accurately for a given turbine only by testing the turMoe for

Sec. 245] STEAM-TURBINE TESTING 269

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

270 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 12

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,

Sec. 246] STEAM-TURBINE TESTING 271

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 Water Rate OjLATurbine 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 sqT 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
section.

248. A Data Form For A Complete Steam Turbine Test is
embodied in the A.S.M.E. Test Code, which is given below:

272 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 12

OBJECT AND PREPARATIONS

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.

APPARATUS AND INSTRUMENTS

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.

{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.

DATA AND RESULTS

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-

Sec. 248] STEAM-TURBINE TESTING 273

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

{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

(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

turbine

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.

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.

18 -^/^ ^'" ' '"'*■ i,--

:STERN UNIVERSITY

\^^/ly ni\/ic\OH

274 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 12

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)

lb.

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) ... .

lb.

(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

(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.

Speed

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.

Sec. 249] STEAM-TURBINE TESTING 275

Power

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)

B.t.u.

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). . .

ft.-lb.
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-

276 STEAM-TURBINE PRINCIPLES AND PRACTICE [Drv. 12

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.

QUESTIONS ON DIVISION 12

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
turbine?

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
condensate.

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.

Sec. 249] STEAM-TURBINE TESTING 277 \

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

DIVISION 13

EFFECT OF STEAM PRESSURE, SUPERHEAT, AND
VACUUM ON STEAM-TURBINE ECONOMY

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

278

Sec. 251] PRESSURE, SUPERHEAT, AND VACUUM

279

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.)

(54)

hiT

Hi — H2
H1-H2

(decimal)

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
cent.

Pres-
sure,
lb. per
SCJ. -in.
gage

Qual-
ity,
per
cent.

Super-
heat,
degrees
Fahr.

Vacuum,

inches

mercury

gage

150

1,110

813

8.58

29.2

150

100

1,195

870

7.84

29.5

8.6

1.0

175

100

1,198

863

7.60

30.3

3.0

2.7

175

150

1,282

917

6.98

30.7

8.1

1.3

175

150

1,282

883

6.38

32.9

8.6

7.1

280 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 13

600

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.)

Sec. 251] PRESSURE, SUPERHEAT, AND VACUUM

281

Fig. 24oJB. — Continuation of Fig. 245J..

282 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 13

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.

Sec. 252] PRESSURE, SUPERHEAT, AND VACUUM 283

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

284 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 13

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

Sec. 255] PRESSURE, SUPERHEAT, AND VACUUM

285

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

versa.)

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.

SUPPLT-STEAM SuPERHEAT.

(Increasing the superheat decreases the
water rate and vice versa.)

Exhaust
Pressure.

Back pressure.
(Increasing back pressure

increases the water rate

and vice versa.)

Vacuum.

(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.

286 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 13

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
below.

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

5400

5200

■55000

O2800

-Am

in2000
fcl&OO
-,1600

;

^1000

-600

400

200

"3!

)

Temperature, Deg.Fahr
Fig. 246. — Graph showing the varia-
tion of the temperature of saturated
steam with the steam pressure.

Sec. 259] PRESSURE, SUPERHEAT, AND VACUUM 287

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

288 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 13

(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.

STEAM CONDITIONS i

!hesi

STAGES

2 3 A

265

68 46 28

4.7

2.4

1.1

LB.ABS0LUTtPRE5SURl

DEGREES SUPERHEAT

'ER TENT MOISTURE

QUALITY OF STtAM

1501 160 1120 175 130

6 1 2 1 4 1 6.5 1 8 1 10.5

Fahrenheit
Thermomefers

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

Sec. 261] PRESSURE, SUPERHEAT, AND VACUUM

289

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

Fig.

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
19

290 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 13

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

Sec. 263] PRESSURE, SUPERHEAT, AND VACUUM 291

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.

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

292 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 13

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
seasons.

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

-Afjlt^ul-t!^

frnrt

1
7 Curve

«n

Z'Vol

ume

..y..

->1

det

»n

1 •

1 '''^

fmosphi

vk Line

Vacuum Llnes^--^

YT-^ ■ ^ . \ 1

^' Zero Pressure Absolute'

1 1 1 1 1 1

...y

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

Sec. 2661 PRESSURE, SUPERHEAT, AND VACUUM

293

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)

294 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 13

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

Sec. 268] PRESSURE, SUPERHEAT, AND VACUUM 295

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.

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.)

296 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 13

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

250.0
50.0

20.2

375.0
75.0

18.3

500.0

Per cent of rated load

100.0

Guaranteed water rate
kw -hr

in lb. per

17.4

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 :

Sec. 269] PRESSURE, SUPERHEAT, AND VACUUM

297

CONDITIONS = 175 lb. per sq.-in. gage; 100° F. superheat; 27-in.

vacuum.

Load in kw

Per cent, of rated load

Water rate in lb. per kw.-hr. by-
acceptance test

500.0
100.0

16.5

The full-load correction graphs (Figs. 252, 253, and 254) for pressure,
superheat, and vacuum corrections, are furnished by the turbine manu-

-r-

26-ln Vaccu

urn

E u

50* F. Superheat.

- 3

—

Full Loaol,500Kw..

-1

>ii

—

5 i

^=V

■)

••

r 2

. .

. _

. -

9'£w

■■

Sj5

[

I€

200

12Q

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-

T-

7ft-ln Vnr

150Lb.Per5c^.ln.&age
Full Loaoi.500 Kw.

^ 3

•|!»

E "rr

vn ^

_ J

5-l5

—

^5=16

—

- -

■

-i

—,

E o

u^-JlC

(

)

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

298 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 13

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 —

"X":

TT-

-r-p

- — 1

DO r. ouperneaT. i

t-- -

150 Lb.fV.r Sn.In CnnpJ

""-■

~||

Full Load. 500 Kw.

n 1 1 III 1 1 1

=16.5---

: -

[JL

"""l

—

E o

«»i

SQ\5

E ^

Oc2

^5j

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

Sec. 270] PRESSURE, SUPERHEAT, AND VACUUM

299

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.-
hr

500.0
100.0

17.1

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 0 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.

300 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 13

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 --------- --^^^^-

!s^

O ' r " "

o^i.

^2+5

■* » S

!s^'iN-_- - - -

^ ---------- - :

_^>.;>^_ - -

- >s^,.

O C A

:. __: !Sfc

-^ i 0 -

:_^h, x- -

^■^

-5^^X-

" " " : - - --"*sS* -

5i ::::::i::::::::::::::::±

:::.:: s ^.

^g.5

1 1 II 1 IJsU

::::::::::::::::::::::j-

^ -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

Sec. 270] PRESSURE, SUPERHEAT, AND VACUUM

301

apparent that the correction of test results to performance guarantee
conditions has led to distortion in listing the performances of actual
installations."

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.
75P00

70.000

65,000

60,000

u 5^000

54000

;^4Woo

i2 40,000

^5i000

^0.000

—

.Tn-h/yl <,-f-enrn Jl-f- IKfi 1 >i

Per 5cf. In. Gage-

•

100° f: Superheat-

Zd" Vacuum

■•

■--N

•>

■

Total

^■hfin

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 0 w 01++5

4.000

4.500

5.000

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,
1922.)

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.

302 STEAM-TURBINE PRINCIPLES AND PRACTICE [Dw. 13

uJKTd+s j-o aniioA uq paiiddv ^g 01 uoipajJOD 4.033 jad

Sec. 270] PRESSURE, SUPERHEAT, AND VACUUM

303

Table 270A. — Guarantee Conditions 200 lb. per sq. in., gage; 100° F.
superheat; 28-in. vacuum.

Load in kw.

2,500

3,750

4,500

5,000

Guaranteed water rate in lb. per

kw,-hr

Total steam — lb. per hr

14.8
37,000

13.9
52,100

13.6
61,100

13.8
69,000

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.

304 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 13

(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

Sec. 270] PRESSURE, SUPERHEAT, AND VACUUM

305

(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

100

34,000

47,600

61,600

15.11

14.10

13.6

+ .85

+ 1.3

+6.77

15.23

14.29

14.49

34,250

48,200

65,100

111

69,000
13.8

+6.77

14.7

73,400

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.,
gage

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.

37,300
14.92

0.85

53,800
14.35

3.2

65,100
14.49

6.77

73,400
14.7

6.77

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

306 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 13

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.

QUESTIONS ON DIVISION 13

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
vacuum?

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.

DIVISION 14
STEAM-TURBINE ECONOMICS AND SELECTION

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-

307

308 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 14

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]

ECONOMICS AND SELECTION

309

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

1,061,000.00

2,775.00

555 . 00

40,600,000.00

7.32

5.23

3.62
1.74

1,210,750.00

2,437.37

322 . 10

35,359,500.00

7.25

4.03

29.20

0.59

0.39

Total operating charges, in dollars, and operating charges per kw.-hr.,

in cents

Total

Per
kw.-hr.

Total

Per
kw.-hr.

Superintendence

Repairs :

Dynamos and appliances ....

Engines

Boilers

Pumps, pipes, fittings, and

miscellaneous

Operating boilers

Operating engines and dynamos

Supplies

Water

Lubricants and waste

Miscellaneous expense

Total, except fuel

Coal

Coal labor, car to boiler room . .

Total cost

Average cost of coal on floor of
boiler room

$122.42

171.33

1,017.48

8.80

880.92

693 . 66

5.47

482.21

220 . 12

291.24

3,893.65

2,635.75

198 . 62

0.014

0.019

0.115

0.001
0.100
0.079

0.055
0.025
0.033
0.441
0.298
0.022

$6,728.02

$1.0214

0.761

$250.10

10.84

299.81

392

390

60.

1,612.

2,177.

114.

15
13
00
80
75
50
08
16
44
62

0.020

0.001

0.024

0.002
0.033
0.032
0.004
0.008
0.004
0.005
0.133
0.180
0.009

$3,904.22
$0.94

0.322

310 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 14

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

32aooo

280,000

Yearly

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:

(57)

^ , - . Average power

Load factor = ^^ — . 1 — — ^

•^ Maximum demand

(decimal)

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

Sec. 276] ECONOMICS AND SELECTION 311

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.

312 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 14

(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
kind.

Sec. 280]

ECONOMICS AND SELECTION

313

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314 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 14

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
indefinite.

pmall lur bines J

^t^

;§

^^^

"■"■/

' inn

^r^^ —

V-

Large Turbine

(About 15.000 /

' ^

•^•in

'.W.)

<->

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]

ECONOMICS AND SELECTION

315

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
Heat

Energy

UNon-

Conden^Ing

Operation

To-tal Energy In A Given Quantity Of 5+eatn Available
For Heating And Power

Power,

Available For Heating

JZCondensinoj
Operation

legends-

Bearing Fricfion And
Radiation

■ 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

316 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 14

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]

ECONOMICS AND SELECTION

317

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.

STEAM TURBINE

STEAM ENGINE

Advantages

Disadvantages

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.

Disadvantages

Advantages

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.

318 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 14

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

—

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Back Pressure = 1.5 Lb. Per 5a. In. Goae

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.5(

100

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
engine.

Sec. 289]

ECONOMICS AND SELECTION

319

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

1.15

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

^11

\
V

.-Compound Con

densing

Corliss Engines

?PM

ide

nsin

9Tl

/rbi'i

les^^

|10

^"^

"~-

iPq

10,1

Horsepower

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.

320 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 14

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.

Turbines

Engines

Condensing Units, Direct-connected

For:
(a) 60-cj/cZe generators in all sizes,
(fe) 25-cycle generators above 1,000-kw.
capacity.

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

essential.

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
engine.

(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
unit.

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

limited.
(6) The water must be recooled and re-
circulated.

Sec. 291]

ECONOMICS AND SELECTION

321

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

— '

•*~

Mr,ri

~f—

?7R

o '^

• — "^^

J »i^

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- t-^T

•»

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^25

"^~

= ! SS«^

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—

= "*

(to

, 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

322 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 14

decrease very much for an increase in capacity. Above
30,000 kw. capacity the price per kilowatt is practically
constant.

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]

ECONOMICS AND SELECTION

323

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.
vacuum.

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

200

Peak load

180

140

1,120

100

500

300

240

Totals.

2,400

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.

324 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 14

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_

Condenser

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]

ECONOMICS AND SELECTION

325

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.

■stack.

■Cjrculafing-water pipes

Fig.

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

326 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 14

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

Kw.-hr.

generated,

from Table I

Steam con-
sumed, lb.

180

23.5

180

4,230

140

24.5

1,120

27,450

100

26.5

500

13,250

29.0

300

8,700

31.0

240

7,440

38.0

2,280

Totals

2,400

63,350

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

ECONOMICS AND SELECTION

327

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

Steam
plant

Oil-engine
plant

Salary, each,
per month

Chief engineer . .
Watch engineers

Oilers

Firemen

175
125
110
110

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

(58)

Unit operating or energy cost

Energy units delivered per year

328 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 14

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

3-100
kw.

Counter-flow
engines

2-200
kw.

a-100

kw.

Investment:

Real estate

Brick building

Generating units, delivered and erected. .
Switchboards and street lighting trans-
formers

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

Fuel

Labor and superintendent

'Lubricants

Miscellaneous supplies

Repairs

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

2,500
35,200
31,954

6,500
3,500
6,500

3,200
13,000
900
1,500
1,800

40,861

2.74

1.93

4.67

562.77

13,000

900

1,500

S 2,500
35,200
25,504

3,500
6,500

4,200
1,800
3,200
13,000

(41,105

2.74

1.95

4.69

570.75

15,915
16,290
8,060

7,000
4,000
6,500

3,200
13,000

$ 2,500
26,000
16,750

3,200
15,400

13,328
20,974
8,060

1.52

5.00

444.25

2,800
1,800
3,200
15,400
900
1,500
1,800

1.62

5.11

474.25

2,500
43,000
23,380

2,500
45,000
24,450

7,000
4,000
6,500

4,200
1,800
3,200
15,400
900
1,500
1,800

3.21

2.03

5.24

591.25

Investment:

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

Fuel

Labor and superintendent. . . ,

Lubricants

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

2,500
35,200
31,954

t 2,500
35,800
32,250

7,000
4,000
7,000

3,900
1,800
3,200
15,100
900
1,500
1,800

t 17,273

12,725

8,060

400

350

1,942

1.97

4.65

575.77

$ 17,513
12,660
8,060

2,500 S 2,500
35,800
25,504 26,634

7,000
4,000
7,000

3,200
15,100

t 42,200
2.96

$ 16,670
14,820
8,060

2.89

1.90

4.79

556.67

2,500
26,000
16,750

28,000
19,449

7,000
4,000
7,000

2,800
1,800
3,200
18,200
900
1,500
1,800

i 14,722
19,900
8,060

$ 2,500
43,000
23,380

3,500
7,000

18,200

900

1,500

S 2,500
45,000
24,450

7,000
4,000
7,000

4,200
1,800

1,500
1,800

S 18,232

17,300

8,060

400

350

1,766

f 46,108

3.18

2.08

6.26

607.75

Sec 293]

ECONOMICS AND SELECTION

329

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

Oil.en

Engines condensing to 26-in. vacuum, turbines condensing to 28-in. vacuum 1

Uniflow engines

Counter-flow
engines

Turbines

Corliss engines
beked

Scmi-diesel

Diesel

a-200

3-100

2-200

3-100

2-200

3-100

2-200

3-100

2-200

3-100

2-200

3-100

kw.

For Satorated Steam

2,500

2.500

2,500

38,000

38,600

38,000

38,600

28,800

30,800

45,800

47,800

26,400

30,100

26,400

23,500

31,954

32,250

25,504

26,634

19,460

23,100

23,380

24,450

56,556

48,220

85,600

72,500

6,500

7,000

6,500

7,000

6,500

7,000

6,500

7,000

6,500

7,000

6,500

7,000

3,500

4,000

3,500

4,000

3,500

4,000

3,500

4,000

3,500

4,000

3,500

4,000

9,500

10,500

9,500

10,500

9,500

10,500

9,500

10,500

2,700

3,000

2,700

3,000

7,000

10,500

7,000

10,500

7,000

10,500

7,000

10,500

4,700

4,400

4., 700

4,400

2,800

3,100

5,000

4,700

5,500

6,000

4,000

4,500

1,800

3,200

1.500

1,500

13,000

15,400

900

1,500

1,800

1,800
.......

1,800

800
2,800
2,500
1,200

800

2,800
2,500
1,200

$125,854

S131,950

1119,404

$126,334

$102,260

$113,700

$127,780

$136,050

$114,256

$111,420

$141,800

$127,600

18,878

19,793

17,910

18,950

15,339

17,055

19,167

20,407

17,138

16,713

'21,270

19,140

11,650

11,900

15,500

15,100

13,100

13,600

17,200

16,700

10,015

7,500

8,060

7,700

400

200

400

438

602

438

602

350

300

350

350-

657

904

657

904

2,138

2,290

1,882

2,065

1,229

1,443

1,891

2,074

2,262
197

1,929
197

3,424
197

2,928
197

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.S8

2.62

2.99

2.96

2.62

2.70

3.19

3.15

2.43

2.27

2.26

2.15

2.26

2.05

2.16

1.75

1.94

2.19

2 33

1.96

1.91

2.44

2.18

4.73

4.88

5.04

5.12

4.37

4.64

5.38

5.48

4.39

4.34

4.71

4.44

629.27

659.75

597.02

631.67

511.30

568.50

638.90

680.25

571.28

557.10

709.00

638.00

For Stteau

Sdperheated 100° F.

% 2,500

S 2,500

$ 2,500

$ 2,500 !$ 2,500

$ 2,500

38,000

38,600

38,000

38,600

28.800

30,800

45,800

47,800

31,954

32,250

25,504

20,634

19,460

23,100

23,380

24,450

6,500

7,000

6,500

7,000

6,500

7,000

6,500

7,000

3,500

4,000

3,500

4,000

3,500

4,000

3,500

4,000

10,000

11,000

10,000

11,000

10.000

11,000

10,000

11,000

7,000

10,500

7,000

10,500

7,000

10,500

7,000

10,500

4,700

4,400

4,700

4,400

2,800

3,100

5,000

4,700

1,800

3,200

15,100

18,200

900

1.500

1,500

1,800

«128,4S4

$134,550

$122,004

$128,934

$104,860

$116,300

$131,080

$139,350

( 19,268

t 20,183

$ 18,300

$ 19,340

$ 15,729

$ 17,445

$ 19,662

$ 20,902

11,100

11,500

14,600

14,700

13,150

14,000

17,350

16,700

8,060

8,000

8,060

400

200

400

350

- 300

300

350

2,222

2,374

1,964

2,149

1,292

1,506

2,003

2,186

% 41,400

S 42,807 [l 43.674

$ 44,999

$ 38,731

$ 41,511

$ 47,825

$ 48,598

2.62

2.59

2.89

2.93

2.63

2.75

3.22

3.15

2.20

2.30

2.09

2.21

1.79

1,99

2.24

2.39

4.72

4.89

4.98

5.14

4.42

4.74

5.46

5.54

642.27

672.75

610.02

644.67

524.30

581.50

655.40

696.75

330 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 14

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]

ECONOMICS AND SELECTION

331

AUIS-CHALMERS MANUFACTURING COMPANY

PROPOSAL

Milwaukee, Wisconsin, U. S. A '^.^.A.

To Brswn...ftna...Bl8.c.lc.M8m£8.e.tuxlng..Cpmpftny.

■ 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).

332 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 14

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.

ALLIS-CHALMERS MANUFACTURING COMPANY,

ACCEPTANCE.

The foregolag proposal i
this day of

(SSK)

hereby accepted
-.J92_

Fig. 272. — Typical manufacturer's proposal (part II; this constitutus pages 3 and 4 of
this particular proposal).

Sec. 293]

ECONOMICS AND SELECTION

333

ALLIS CHALMERS MAN LFACTL RING COMPANY

MILWAUIBB. WISCONSIN, U S. A.

These guai

MIXED PRESSURE CONDENSING STEAM TURBINE
MIXED rh. ^^^ ALTERNATOR UNIT

Brow^ and BlaoH Manufacturing Company. St. Louis. Mo.

and specifications (orm part of proposal dated J*"®.. J '

STEAM CONSUMPTIONS

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. hour...ie. 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&

Ther

. factor 750 K. W. Maximum.

Amperes.

Phase_ _..a :; -

revolutions per i

IC W

7»

780 BQ - ^^'''^

Excitation voltage 125 Appro:

required with rated curre

for

Insulation

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.

0 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

feet.
46&00

:t.ng 3£360
erection 3180

pounds,
pounds.

lount ol local COI
(NOTE— U more loom !• no
I pr^edence over Ihf fnnui t

„SiS.V

Kunes hanwna I
.ith sublrtteri. 1

FiQ. 273.— Typical manufacturer's proposal (part III; this constitutes pages 5 and 6 of
this particular proposal).

334 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 14

UIXED PRESSURE CONDENSING STEAM TURBINE AND ALTERATOR UNIT.

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]

ECONOMICS AND SELECTION

335

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

MIXED PRESSURE CONDENSING STEAM TURBINE AND ALTERATOR UNIT.

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.

ALJ.IS-CHALMERS MANUFACTURING COMPANY,

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

336 STEAM-TURBINE PRINCIPLES AND PRACTICE [Div. 14

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.

QUESTIONS ON DIVISION 14

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?
Why?

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?
Why?

Sec. 294 ECONOMICS AND SELECTION 337

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?

SOLUTIONS TO PROBLEMS ON DIVISION 1
STEAM-TURBINE FUNDAMENTAL PRINCIPLES

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
ft.-lb.

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.

SOLUTIONS TO PROBLEMS ON DIVISION 8
REGENERATORS AND CONDENSERS

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.

338 STEAM-TURBINE PRINCIPLES AND PRACTICE [Drv. 14

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.

SOLUTIONS TO PROBLEMS ON DIVISION 9

HIGH-PRESSURE, BLEEDER, MIXED-PRESSURE, AND
EXHAUST-STEAM TURBINES

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.
(30):

^^^ =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.

INDEX

Acceptance test, water rate correction, 295
Accumulator, see Regenerator .
Adjustment, see also Clearance, Align-
ment.

axial, rotor, 100

speed, see Governor.
Air leak, 244

Air-pressure governor, Ridgway, 148
Alberger Pump & Condenser Co., 69
Alignment, see also Clearance, Adjust-
ment.

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,

34
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
Gurney,98
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

339

340

INDEX

Bucket-wheel turbine, 24

definition, 38

illustration, 28
Bypass governor, see Governor,

valve, see Valve.

Capacity, generating, eflfect of load factor,
311

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-
ment.

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

TURBINE, 34

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

ALIGNMENT, 171

Sturtevant, 168
care, 170
FLEXIBLE, 167

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

TYPE TURBINE, 59

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

INDEX

341

Diaphragm, 56

Direct-current generator, starting,
246
stopping, 246
Disc, 25
Double reduction gears, see Reduction

gears.
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

turbine.
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-
ing.
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,
310

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,
136
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

342

INDEX

Governor, vacuum breaker, 125
VALVE, definition, 155

low-pressure turbine, transverse sec-
tion, 216
mixed-pressure turbine, 199
Terry turbine, 127
Westinghouse centrifugal, adjustment,
139
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
turbine.

INDEX

343

Mixed-pressure turbine, construction,
197
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-

gency.
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,
110

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,
78

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

344

INDEX

Quotation, requesting, 330
R

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,
49
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,
279
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

INDEX

345

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,
219

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,

286
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,
295
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,

17
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,
278-306
relative to engine economy, 316
efficiency, 314
efficiency ratio, 17
energy losses, 17

346

INDEX

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-
bine.
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

turbine.
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
Vacuum

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,

325
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

INDEX

347

WestinGhottse Electric & Mfg. Co.,
bleeder turbine, 191
coupling, 169
emergency governor, 142
expansion joint, 183
geared-turbine and generator, 165
GOVERNOR, 136

adjustment, 139
impulse turbine, 48, 84

TURBINE nozzles, 50

IMPULSE-AND-REACTION TURBINE, dou-

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

WORTHINGTON PuMP AND MACHINERY

Corp., condenser graph, 181
installation with barometric-jet con-
denser, 178

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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)

12962

;

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