Power Plant Familiar is at Ion v- III

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TYPES AND WORKING PRINCIPLES STEAM TURBINES

1.0 INTRODUCTION

Steam turbine is a rotating machine which- CONVERTS HEAT ENERGY OF

STEAM TO MECHANICAL ENERGY.

In India, steam turbines of different capacities, varying from 15 MW to 500

MW, are employed in the field of thermal power generation. The design, material,

auxiliary systems etc. vary widely from each other depending on the capacity and

manufacturer of the sets. Therefore the discussions in the chapters will follow the

general patterns applicable to almost all types of turbines, with reference to the

specific features of 21 0 MW steam turbines (both L.M.W. Soviet & KWU German

Designs)'and 500 MW (KWU) turbines which form the backbone of the thermal

power sector in India.

1.1 DEVELOPMENT OF STEAM TURBINE

Historically, first steam turbine was produced by Hero, a Greek Philosopher, in 120

B.C. (Fig. 1. l.). As the fig. shows, k was a pure reaction turbine (explained at 1.4).

In 1629, an Mlian. named Branc actually anticipated the boiler-steam turbine

combination that is a major source of power today. The concept, is illustrated in

(Fig. 1.2).

First practical steam turbine was introduced by Charles Parsons in 1884

which was also of the reaction type. Just after five years, in 1889, Gustav De Lava]

produced the first practical impulse turbine.

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Active development of steam turbine made ft the principal prime mover of

generating stations by 1920. Most units used 14 kg/cm2 and 276oc steam and

capacity ranged from 5,000 'La 30,000 KW. By 1930 steam M2 conditions rose to

48 kg/c and 398oc and by 1940 steam condition of 81 kg/cm' and 509oc was

achieved.

After second world war (1 945), reheat. cycle was adopted widely and

capacity increased gradually. While turbines of 900 MW are in use in USSR, in

India the largest capacity is 50&MW with steam condition of 179

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1.2 WORKING PRINCIPLES

When steam is allowed to expand through a narrow orifice, ft assumes kinetic

energy at the expense of its enthalpy (heat emrgy). This kinetic energy of steam is

changed to mechanical (rotational) energy through the impact (impulse) 6r reaction

of steam against the blades.

It should be realized that the blade of the turbine obtains no motive force

from the static pressure of the steam or from any impact of the steam jet. The

blades are designed in such a way, that steam will glide on and off the blade without

any tendenc to strike it.

As the steam moves over the blades, its direction is continuously changing

and centrifugal pressure exerted as the result is normal to the blade surface at all

points. The total motive force acting on the blade is thus the resultant of all the

centrifugal forces plus the change of momentum. (Fig. 1.3). This causes the

rotational motion of the blades.



FIG. 1.4 SIMPLE IMPULSE TURBINE

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1.3 TURBINE TYPES

Basically there are two broad classifications of steam turbines:

i) Impulse: In Impulse turbine(Fig.1.4),the steam is expanded (i.e. pressure is

reduced) in fixed nozzles. The high-velocity steam issuing from the nozzles does work

on the moving blades which causes the shaft to rotate, The essential feature of an impulse

turbine is that all the pressure drops occur in the nozzles only, and there is no pressure

drop over the moving blades.

ii) Impulse-reaction : In this type, pressure is reduced in both fixed and moving

blades. Both fixed and moving blades act like nozzles and are of same shape. Work is

done by the impulse affect due to the reversal of direction of the high velocity steam plus

a reaction effect due to the expansion of steam through the moving blades. This turbine

is commonly called a reaction turbine and is shown below in (Fig.1.5).



1.4 COMPOUNDING

Several problems crop up if the energy of steam is converted in one step, i.e. in a

single row of nozzle-blade combination. With all heat drop taking place in one row

of nozzles (or single row of nozzles and blades in case of reaction turbine) the

steam velocity becomes very high and even supersonic (velocity of steam is

proportional to square root of heat drop in nozzle; V=44.8/K(H1-H2) m/s.

K=constant, H, Enthalpy at nozzle inlet ; H 2 Enthalpy at nozzle outlet. The

rotational speed of the turbine also becomes very high and impracticable.

So, in order to convert the energy of steam within practical speed range, it is

necessary to convert R in several steps and thus reducing the velocity of steam and

rotor speed to practical levels. This is termed compounding.

Following are the various types of compounding.

1.4.1 Velocity Compounded Impulse Turbine (Fig. 1.6)

Like simple impulse turbine this has also only one set of nozzles and entire steam

pressure drop takes place there. The kinetic energy of high velocity steam issuing

from nozzles is utilised in a number of moving row of blades with fixed blades in

between them (instead of a single row of moving blades in simple impulse turbine).



The role of the fixed guide blades is just to change the direction of steam jet and

guide it to next row of moving blades. This type of turbine is also called curtis

turbine.

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1.4.2 Pressure Compounded Impulse Turbine (Fig.1.7)

This is basically a number of simple impulse turbines in series on the same shaft -

the exhaust of one steam turbine entering the nozzle of the next turbine. The total

pressure drop of the steam does not take place in the first nozzle ring, but is

divided equally between all of them. Steam is passed through the first nozzle ring

in which it is only partially expanded. It then passes over the first moving blades

wheel where most of its velocity is absorbed, From this ring it exhausts into the

next nozzle ring and is again partially expanded. The velocity obtained from the

second nozzle ring is absorbed by the next wheel of moving blades. This process

is repeated in the remaining rings until the whole of the pressure has been

absorbed. This type of turbine is also called Rateau turbine after its inventor.

1.4.3 Pressure-Velocity Compounded Impulse Turbine (Fig. 1.8)

Pressure-Velocity Compounding is a combination of both the previous methods

and has the advantage of allowing a bigger pressure drop in each stage and so less

stages are necessary. Hence, for a given pressure drop the turbine will be shorter.

But the diameter of the turbine is increased at each stage to allow for the



increasing volume of steam. This type was once very popular. But it is rarely,

used as efficiency is quite low.

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1.4.4 Multistage Reaction Turbine

(Fig. 1.9) shows such a multisage reaction turbine consisting of a number of rows

of moving blades attached to the rotor and an equal number of rows of fixed

blades attached to the casing. Each stage txilises a portion of energy of steam.

Theoretically this may be called pressure compounded turbine as the pressure of

steam drops gradually over the succeeding stages.

FIG. 1.8 COMPOUDING FOR PRESSURE FIG.1.9 REACTION TURBINE PRESSURE AND

AND VELOCITY VELOCITY CURVES

The fixed blades compare to the nozzle used in the impulse turbine. Steam is

admitted over the whole circumference, and in passing through the first row of fixed

blades, undergoes a small drop in pressure and its velocity is increased. It then

enters the first row of moving blades and, as in the impulse turtiine, suffers a

change in direction and hence momentum giving an impulse on the blades. During

the steam passage through the moving blades, k undergoes a further small drop in

pressure resulting in an increase in velocity which gives riseto areaction inthe

direction opposite tothat of the added velocity. It is in thisthat the impulsereaction

turbine differs from the pure impulse turbine. Thus the gross propelling force in the

impulse-reaction turbine, (or the 'reaction' turbine as it is commonly called), is the

vector sum of the impulse and the reaction forces.

(Fig. 1.9) also shows how the blade heights increase as the specific volume of

the steam increases with reduction in pressure. Note, how the pressure falls



gradually as the steam passes through the groups of blades. There is a pressure drop

across each row of blades both fixed and moving. This is of considerable practical i

, especially atthe high pressure end of the turbine where the pressure drops are

greatest. Because this diftererwe of pressure tends to force some steam through the

clearance spaces between the nxmng b~s and the casing and between the fixed

blades and the rotor. These dearances have to be carefully controkd by using axial

and 1 or radial seals at the blade tips, otherwise the leakage would be so large that

the turbine would be ineffi~t. The pressure drop across the moving blades gives

rise to a large axiaj thrust on #m rotor, towards the low pressure end of the turbine,

and special balance pistons / thrust

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bearings have to be fitted to counteract it. The dummy (balance) piston diameter is

so calculated that the steam pressure acting upon it in the opposite direction to the

steam flow, balances out the force on the rotor blades in the direction of steamflow.

Preferably the dimensions are so arranged to keep a small thrust towards the inlet

end of the turbine. To maintain this condition at all loads in some designs, a

balance pipe is usually connected from the casing, on the outer side of the balance

piston, to some tap off point down the cylinder. This pipe-maintains the steam

pressure on the out board side of the dummy pistonto correspond with pressure at

the stage down the turbine cylinder to which the balance pipe is connected. Under

steady load conditions the steam leakage through the dunvny piston "rinth packings

flows, from out board side of dummy piston, through the balance pipe and does

workin the lower stages of turbine. This arangement is shown in (Fig.1.10).



The steam velocities in this type of turbines are moderate. The velocity of

steam for maximum blade efficiency being roughly equal to the blade velocity. The

leaving loss is normally about the same as for the muffi@age impulse turbine.

The impulse-reaction turbine was developed by the late Sir. Charles A.

Parson and widely used in power stations. It is sometimes called Parson's turbine.

1.5 IMPULSE VS REACTION-PRESENT TREND

The hard and fast distinction between the impulse reaction is becoming

progressively less important. The trend is to have some percentage of reaction for

an impulse turbine or to have some percentage of impulse for a reaction turbine.

It can be mathematically provedthat efficienc@of reaction stage is

greaterthan efficiency of impulse stage. A pressure difference exists across the

reaction type moving blades, therefore, the changes of leakage of steam from

around the blade is more in a reaction stage. The advantage of efficiency is off set

by the inter stage leakage of steam which flows without doing useful work. Hence

a reaction stage should be located in the low pressure region of turbine.



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There is a general rule to use a greater percentage of impulse on the HP end

and greater percentage of reaction on the L.P. end. The percentage of reaction

progressively increases as we go towards L.P. end.

In actual turbines it is common for the best feature of various types to be

incorporated in one machine. Forexample, a turbine may have avelocity

compounded (curtis) firststagefollowed by pressure compounded impulse (Rateau)

stages and, at the low pressure end of the machine, reaction blading.

1.6 CLASSIFICATION OF STEAM TURBINES

Steam turbines may be classified into different categories depending on their

construction, the progress by which heat drop is achieved, the initial and final

conditions of steam used and their industrial usage etc.

1.6.1 According to the Directon of Steam Flow

a) Axial turbines:in which the steam flows in a direction parallel to the axis of the

turbine.

b) Radial turbines: in which the steam flows in a direction perpendicular to the

axis of the turbine.

1.6.2 According to the Number of Cylinders

a) Single - cylinder turbines

b) Double - cylinder turbines

c) Three - cylinder turbines and

d) Four - cylinder turbines etc.

1.6.3 According to the Method of Governing

a) Turbines with throttle governing in which fresh steam enters through one or

more (depending on the power developed) simultaneously operated throttle valves.

b) Turbines with nozzle governing which fresh steam enters through two or

more consecutively opening regulators.

c) Turbines with by-pass governing in which steam besides being fed to the

first stage is also directly led to one, tow or even three intermediate stages of

turbine.

1.6.4 According to the Principle of Action of Steam

a) Impulse Turbine



b) Reaction Turbine

1.6.5 According to the Heat Balance Arrangements

a) Condesnsing turbines with regeneration; in these turbines steam ad a

pressure less than atmospheric is directed to a condenser; besides, steam is

also extracted from intermediate stages for feed water heating, the number

of such extractions usually varies from 2-3 to as much as 8-9. Small

capacity turbines of earlier designs often do not have regenerative feed

heating.

b) Condensing turbines wfthoneortwoin termediate stage extractions at

sp'opific pressures for in dustrialand heating purposes.

c) Back pressure turbines : the exhaust steam from these turbines is utilised

for industrial or heating purposes.

d) Topping turbines: these turbines are also of the back pressure type with the

difference that the exhaust steam from these turbines is furtherutilised in

mechum and fow-Kessure condensing turbines. These turbines, in general,

operate at high initial conditions of steam pressure and temperature, and are

mostly used during extension of power station capacities, with a view to

obtain better effidencies.

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By extension of power stations capacities here is meant additional

installation of high pressure boiler (critical and super critical pressures) and

topping turbines as additional units, del~ng steam to the already existing

medium-pressure turbines from the exhaust of topping turbines.

e) Back pressure turbines with steam extraction from in terme diate stagesat

specific pressures ;turbines of this type are meant for supplying the

consumer with steam of various pressure and temperature conditions.

f) Low-pressure (exhaust-pressure)turbine sin which the (exhaust-steam) from

reciprocating steam engines, power hammers, presses etc. is utilised for

power generation purposes. g) MLxed pressure turbines with two or three

pressure stages, with supply of exhaust steam to its intermediate stages.

1.6.6 According to the Steam Conditions at Inlet to Turbines

a) Low-pressure turbines, using steam at pressure of 1.2 to 2 ata.



b) Medium-pressure turbines, using steam at pressure of up to 40 ata.

c) High-pressure turbines, utilising steam at pressures above 40 ata. and below

170 ata.

d) Turbines of very high pressures, utilising steam at pressures of 170 ata and

higher and temperatures of 550oc and higher.

e) Turbines of superc~i pressures, using steam at pressures of 225 ata and

above.

1.6.7 According to Shaft Arrangements

i) Tandem compounded turbines - Here all the cylinders are arranged so as to

drive a single shaft.

ii) Cross compounded turbines - Here cylinders are arranged to drive two or

more shafts with separate generators with every shaft.

1.6.8 Automatic 1 Non-Autoniatic Extraction Turbines

Automatic-extraction unit bleeds off part of main steam flow at one, two or three

points. Valved partitions between selected turbine stages control extracted steam

pressure at the desired level.

Non autom@-extraction turbines bleed steam at as many as nine different

stages. Pressure of extracted steam at each state varies with the turbine shaft load;

extracted steam is used for feed heating.

(Fig. 1. 1 1) & (Fig. 1. 1 2) show various types of turbines.

1.7 BASIC PRINCIPLES

The Thermal Power Plants with steam turbine uses Rankine cycle. Rankine cycle

is a vapour power cycle having two basic characteristics:

i) the working fluid is a condensable vapour which is in liquid phase during

part of the cycle and

ii) the cycle consists of a succession of steady flow processes, with each process

carded out in a separate

component specially designed for the purpose. Each constitute an open

system, and all the components are connected in series so that as the fluid

circulates through the power plant each fluid element passes through a cycle

of mechanical and thermodynamic stages.

1.7.1 Temperature Entropy Diagram (Discussed in Detail in Vol.1)



The teniperature-entropy (T-S) diagram is probably the most useful diagram of all

illustrating certain fundamental points about Rankine steam cycles. Ideal

condition for a unit on a T.S. diagram are indicated in

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(Fig. 1. 1 3). The unit uses steam at a pressure of 1 00 bar absolute, temp. 540'c

(813'K) and rejects it to the condenser at 30 m bar (saturation temp. 24.1 "c).

At point'A'the condensate is at boiling temperature corresponding to the back

(condenser) pressure. Its pressure is raised to 1 00 BAR in Feed Pump

corresponding to point'B'. Heat (sensible) is added to this water to raise its

temperature. At the point C it reaches its saturation temp. at a pressure of 1 00 bar.

Evaporation begins at the point C. Heat (latent-because no rise in temperature

between C and D, as evident from the diagram), addition continues. At D all the

water evaporates and super-heating commences. This is shown by the curve DE.



Steam then expands isentropically i.e. enters the turbine and rotates ft, as

shown by the line EFG. At point F there is not superheat left in the steam and so

from F to G there is increasing wetness. At G steam is at a pressure of 30m bar and

is passed out of the turbine to the condenser and condensation of steam takes place

as represented by the line GA. At point A the steam has all been condensed and

condensate is at boiling temperature ready to begin another cycle.

To summarise the above:

AB pressure Rise in BFpp.

BC heating of feed water (i.e. sensible heat addition)

CD evaporation of water in boiler (i.e. latent heat addition)

DE superheating of steam (i.e.superheat addition)

EFG expansion of steam in turbine, point E denotes demarcation between

superheated and wet steam.

GA condensation of steam in the condenser.

An important basic fact to remember is that heat is product of absolute

temperature and change of entropy. In other words heat is represented by the area

under the diagram.

FIG. 1.13 SENSIBLE LATENT AND SUPERHEAT, 100 BAR, 540OC CYCLE ON T-S DIAGRAM



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1.7.2 Velocity Diagrams

Let us consider an axial-flow turbine consisting of one or more stages, each stage

comprising one annulus of fixed nozzles and one annulus of moving blades as in

(Fig. 1.14) (a). Usually the total pressure drop across the stage is divided between

the nozzle and blades as illustrated, (Fig. 1.14) (b). The division is usually

expressed, not in terms of pressure drops, but in terms of the corresponding

enthalpy drops. The criterion used is the degree of reaction A, defined as

A = Enthalpy drop in moving blades (h1-h2)

Enthalpy drop in stage (h0-h2)

In Impulse Turbine, A < 0.5

In Reaction Turbine, A >0.5

The mode of action of the turbine can best be studied by following the path of

fluid through a single stage at the mean radius of the annulus, as in (Fig. 1.15)

(a). The fluid enters the nozzles with velocity C0 at pressure Po and is expanded

to pressure P0 . It leaves the nozzles with a velocity C1, in a direction making an



angle α1 with the tangential direction, i.e. the plane of rotation, It must satisfy

the energy equation.

½ (C12 – C0

2) = h0 – h1

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The velocity of the fluid relative to the moving blades can be found by

subtracting vectorially the blade speed U. This is easily accomplished by drawing

the inlet velocity triangle. To avoid a multiplicity of indices, relative velocities

are denoted by V, and the relative velocity at inlet to the blades is therefore V,,

V, makes an angle P, with the tangential direction, and if the fluid is to flow

smoothly into the blade passages without undue disturbance, the inlet angle of

the blades must be made approximately equal to P, If the outlet angle of the rotor

blades is P2'the direction of the relative velocity at outlet, V2 will also be

approximately 02. Applying the energy equation to the flow relative to the rotor

blades it follows that

1/2 (V22 - V1

2) = h1 - h2

Since no work is done relative to the blades, i.e. in the frame of reference of the

blades, V2 must satisfy this equation. Vectorial addition of the blade speed then

givesthe absolutevelocity C2 andthe direction CL2; Usually, if this is one stage

of a multistage turbine, C2 will be made equal to Cl and ct 2 equal to a.; the fluid

can then pass on to another similar stage.

1.7.3 Work done in a stage

Equation below gives the work done i.e.

W = mU. (C1w - C2w) = mU Cw

Where Cw denotes the tangential component of fluid velocity or commonly

called the 'whirl velocity'.

An alternative expression for the work done in terms of the fluid velocities is

given by:-

W = m/2 (C12 _ C2

2) + (V22 _ V1

2)

WORK DONE IS THE KINETIC ENERGY AT INLET TO THE ROTOR

BLADES mC,212, PLUS THE ENERGY M(V22 _ V,2) MADE AVAILABLE BY



EXPANSION IN THE ROTOR MINUS THE REJECTED KINETIC ENERGY

mC22

/ 2. Since in an impulse stage turbineV2 = V1

W = 1 / 2m (C1 - C2 ) 2

1.7.4 Turbine Losses

The losses in a turbine may be divided conveniently into two groups external

and internal.

There will be external losses due to bearing friction and the power required to

drive auxiliaries. These losses will be similar for all types of turbine and will not

be considered here.

The second and major group, the internal losses, are basically of three types;

a) Friction losses: The fluid friction in the nozzle and rotor blade passages

results in the actual enthalpy drop being less than the isentropic enthalpy

drop. The losses due to this are friction losses.

b) Leakage losses : Because clearance is needed between the moving and

stationary parts, some fluid passes through the turbine without doing its full

complement of work on the blading. Losses incurred in this way are called

leakage losses.

On modern turbines inter-stage leakage accounts for 0.5 to 1.0% loss if the

seals are in good condition.

c) Leavingloss: The considerable kinetic energy of steam when ftieaves the last

row of moving biadesdoes not do further useful work. This loss of energy is

known as 'Leaving Loss' or "residual loss";

Thus, Leaving loss = MVO2 1 2 where M = Mass steam flow, V. = absolute

velocity of steam at the outlet of last row of moving blades.

This loss varies as square of the velocity of steam. To minimize this loss R is

important to keep the velocity of steam leaving the last wheel as low as possible.

To achieve this the annular area (i.e., blade height x mean diameter) of last row of

blading is made as large as practicable. This is done by increasing the height of last

stage blades.

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TURBINE DESCRIPTION

2.1 GENERAL DESCRIPTION OF A 210 MW (LMW) STEAM TURBINE



Before discussing in details about various features of the steam turbine and its

auxiliaries let us have an over view of the system as a whole. (Fig.2.1) shows the

schematicofa21OMW steam turbine (BHEL/LMW).

Superheated steam (130KglcM2,535'c) from the boiler enters in to the High

pressure turbine through two emergency stop valves (ESVS) and four control valves

(CVs).

The high pressure turbine (HPI) comprises of 12 stages, the first stage being

governing stage. The stean flow in High pressure turbine (HPT) being in reverse

direction, the blades in high pressure turbine HPT are designed for anticlockwise

rotation, when viewed in the direction of steam flow.

After passing through High pressure turbine (HP]) steam (27 Kg/cm 327'C)

flows to boiler for reheating and reheated steam (24.5 Kglcm2, 535oC) comes to the

intermediate pressure turbine (IP'T) through two interceptor valves (IVs) and four

control valves (CVs) mounted on the IPT it self.

The intermediate pressure turbine has 1 1 stages. High pressure turbine

(HPT) and intermediate pressure turbine (I PT) rotors are connected by rigid

coupling and have a common bearing.

After flowing through intermediate pressure turbine (IPT), steam enters the

middle part of low pressure turbine (LP]) through two crossover pipes. In low

pressure turbine, steam flows in the opposite paths having four stages in each path.

After leaving the low pressure turbine the exhaust steam (0.09 KgIcm2 abs)

condenses in the condensers welded directly to the exhaust part of the low pressure

turbine.

Rotors of intermediate and low pressure turbiners are connected by a semi

flexible coupling.

The direction of rotation of the rotors is clock wise when viewed from the

front bearing end towards the generator. The three rotors are supported in five

bearings. The common bearing of High pressure and Intermediate pressure rotors is

a combined journal and radial thrust bearing. .

Turbine is equipped with a turming (barring) gear which rotates the rotor of

the turbine at a speed of nearly 3.4 rpm for providing uniform heating during

starting and uniform coiling during shut down. Seven steam extractions for feed



water heating have been taken from gth, 12th, 1 Sth, 18th, 21 st, 23rd & 25th

stage.Condensate from the hot well of condenser is pumped by the condensate

pumps, and supplied to the deaerator through ejectors, gland steam cooler and four

number low pressure heaters. Steam is extracted from the various points of the

turbine to heat the condensate in these heat exchangers. From deaeratc)r the feed

water is supplied to boiler by boiler feed pumps through three number High

pressure heaters.

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2.2 200 / 210 MW (KWU) STEAM TURBINE

(Fig.2.2) shows the sectional arrangement of a 21O MW steam turbine of BHEL /

KWU make.21OMW KWU turbine is a Tandem compounded, three cylinder,

single reheat, condensing turbine provided entirely with reaction blading.

Superheated steam (1 47 Kg/ cM2 abs, 535OC) enters the HP turbine through two

initial steam stop and control valves. HP cylinder has a throttle control. HP turbine

comprises of 25 single flow stages. HPTexhauststeam (39.2 kg/cm', 343'c)

goesforreheating and reheated steam (34 kg /CM2 535'c) comes to

intermediate pressure turbine through two combined reheat stop and control valves.

The lines, leading from the two HP exhaust branches to Reheater, are provided with

swing check valves which prevent hot steam from the reheater flowing back into the

HP turbine. IPT is a double flow turbine with 20 reaction stages per flow. From 1

PT, steam goes to double flow LP turbine with eight reaction stages per flow

through cross around pipes. The steam from LPT is exhausted to condenser at a

back pressure of 0.1 187 bar (49'c). Extraction steam is bled from six bleeding

points (3-LP Heaters, 2-HP heaters, one deaerator). The individual turbine rotors

and the generator rotor are connected by rigid couplings.

2.3 500 MW (KMU) UNIT

Like 200 MW KWU design, 500 mw turbine is a Tandem compounded, three

cylinder, single reheat condensing turbine provided entirely with reaction blading.

Superheated steam (1 70 Kg / CM2 , 537'C) enters the H.P turbine through four



combined main stop & control valves. HP turbine comprises of 18 single flow

stages. HP turbine exhaust steam (45 kg / CM2 , 342.5'c) goes for reheating and

reheated steam (40.5 kg cm2, 5370c) comes to I P turbine through four combined

reheat stop and control valves.

I.P. turbine is a double flow turbine with 14 stages per flow. The LP turbine is

a double flow turbine with 6 stages per flow. Condenser is maintained at a preasure

of 0. 09 kg / CM2 (abs). Extraction steam is bled from six bleeding points.

The turbine shaft layout is similar to that of 21 0 MW KWU design. As in 21

0 MW unit, the HP rotor is supported by two bearings, a journa[Dearing at the front

end of the turbine and a combined journal and thrust bearing directly adjacent to the

coupling with the ]P rotor. The iP & LP rotors have a journal bearing each at the

end of the shaft. The bearing temperatures are measured by thermocouples in the

lower shell directly under the white metal lining. The temperature of the thrust

bearing is measured in two opposite thrust pads.

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TURBINE COMPONENTS

3.1 CASINGS OR CYLINDERS

A casing is essentially a pressure vessel which must be capable of withstanding the

maximum working pressure and temperature that can be produced within it. The

cylinder is supported at each end. The cylinder has to be extremely stiff in a

longitudinal direction in order to prevent bending and to allow accurate clearances

to be maintained between the fixed and moving parts of the turbine. This

determines the length between bearing centres which in turn determines the number

of stages which can be accommodated within the cylinder.

The working pressure aspects demand thicker and thicker casing and the

temperature aspects demand thinner and thinner casings. Design developments

took place to take care of both pressure and temperature considerations and resulted

in the following three types of casing design.

i) Single shell casing

ii) Multiple(double)shell.easing

iii) Barrel type casing

3.1.1 H.P. Turbine Casing



a) Single Shell Split Casing : Barlier design turbines including the 21O MW

BHEL / L.M.W. varieties are of single shell split casing for H.P. cylinders.

In this type the casing thickness would be of the order of about20 cms for

the 21 0 MW turbine which will make the flange to about 40cms and the

jointing bolts to about 23 crm size. This ' leads to concentration of mass

where high temperature and sharp fluctuation in temperature h expected.

This poses several problems during machine start ups and load changes.

b) Double Sheli Casinci: With the rise of steam conditions there fore single

shell casings are of no more use for H igh Pressure (H P) and 1 ntermediate

Pressure (1 P) casings. By using a double shell casing , the casing

thickness has been reduced to 9 cms and bolt size to 1 1 cms. in 21 0 MW

turbine H.P. cylinder. (Fig. 3. 1) shows how a double shell reduces

temperature difference through metal casing.

FIG. 3.1 SHOWING HOW A DOUBLE WALL CYLINDER REDUCES

TEMPERATURE DIFFERENCE THROUGH METAL CASING

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In turbines where steam temperature are veryhigh, the HP cylinder is

generally of the double shell design, the inner shell carries the stationary blading

and diaphragms and is subject to full steam pressure, whilst the space between the

two shells is subjected to the exhaust steam pressure of the particular H.P.

cylinder. The advantage of this arrangement is that each shell need only be

designed for a relatively small pressure difference. This permits reduced shell

thickness and allows quicker warming up without undue stress when starting up .



Special expansion joints are provided for the main steam inlet pipes to pass

through the outer shell and connect on to inner shell. (Fig.3.2) illustrates,a typical

double shell H.P. cylinder. (Fig.3.3) shows longitudinal section of a H.P. turbine

with double shell casing.

BHEL 21 0 mw Turbines of Soviet Design have single casing type HP

cylinders, while 21 0 MW BHEL KWU Turbines have barrel type for HP cylinder.

PAGE 21

c) Barrel Type Casing : The barrel type of cylinder construction ensures

symmetry of the wall thickness around the aixs of rotation and hence the wall



thickness itself is relatively less than that used in other type of constructions.

Barrel Design has also been used successfully for other turbo machines working at

high pressure, e.g. Feed Pump. Because of its rotational symmetry, the barrel type

casing also remains constant in shape and leak proof during quick changes in

temperature (e.g. on start-up and shutdown, on load change and under high

pressures).

The principal parts of the HP Turbine casing are an axially split inner shell

enclosing the rotor and an outer shell of barrel type casing (Fig.3.4). (Fig. 3.5)

shows how the inner shell is accommodated into the barrel.

The space between inner shell and the barrel type casing is sealed from the

neighboring spaces by a sealing with U &1 cross-section. The axially splitinner

shell is constructed as a guide blade carrier. The location of guide blade carrier

within outer casing enables, independent expansion of each other, i.e., it is radial

in all directions and axial from a fixed point.

PAGE 22

The guide blade carrier is kinematrically supported and located and this is

guaranted by 4 projections on e,@ s the inlet side and 2 projections on the exhaust

side of the guide blade carrier, which match with corresponding e ofslotsand

havesliding contacts. Thesliding contacts make it possibleforthe blade carrierto be

exactly guided igh into the vertical and horizontal planes. A ring with a buttress

thread holds the guide blade carrier in the barrel-ant type casing. As the



combined journal and thrust bearing on the inlet side of the turbine forms the fixed

pointge of the turbine shaft, the guide blade carrier and turbine shaft expand in the

same direction. This means that only small aaxial clearances between the two parts

are necessary.

(Fig. 3.5) shows the 500 MW (KWU Design) HP Turbine with barrel type

casing assembly.

PAGE 23

FIG. 3.5 ASSEMBLY SEQUENCE FOR AN HMN – RANGE H.P. TURBINE

Steam Admission

Steam is admitted to HP turbine either through two or four branches as per

designers choice.

In 210 MW BHEL_LMW design steam is admitted through four branches, each having a

control valve (CV) in it.



PAGE 24

In 210 MW BHEL-KWU Design, steam is admitted through two branches each having a

combined stop valve and control valve in it.

The following (Fig.3.6) shows different versions of steam admision to HP turbines.

FIG. 3.6 DIFFERNT VERSIONS OF THE h.p. TURBINES

3.1.2 I.P. Turbine Casing

In modern large capacity turbines the casing of the IP turbine is split horizontally

and is of double-shell construction. (Fig. 2.2) shows a double-flow type I.P.

turbine casing (KWU-210 MW). Steam from the HP turbine enters the inner

casing from above and below through two inlet nozzles flanged to the mid section

of the outer casing. This arrangement provides oppesed doublke flow in two blade

section and compensates axial thrust.

In BHEL / LMW 210 MW set the IP turbine is of single flow type (see Fig.

2.1). The casing here is made of two parts, The two parts are connected by a



vertical joint. Each part consist of two halves having a horizontal joint. The

horizontal joint is secured with the help of studs and nuts.

From I.P. turbine steam is carried through two cross-over pipes to the

double flow L. P. cylinder. Each cross-over pipe is provided with a compensator

for taking care of thermal expansion and to ensure that no

PAGE 25

FIG. 3.7 I.P. TURBINE OF 500 MW TURBINE

heavy thrust or turning moments are thrown on to the I.P. cylinder. (Fig. 3.7)

shows cut away section of 500 MW Unit I.P. turbine.

3.1.3 L.P. Turbine Casing

The LP turbine casing shown in Fig. 2.2 consists of a double-flow unit and has a

triple shell welded casing, The outer casing consists of the front and rear walls, the

two lateral longitudinal support beams and the upper part. The front and rear

walls, as well as the connection areas of the upper part are reinforced by means of

circular box beams. The outer casing is supported by the ends of the longitudinal

beams on the base plates 3.2 of the foundation.



The double-flow inner casing, which is of double-shell construction,

consists of the outer shell and the inner shell. The inner shell is attached in the

outer shell with provision forfree thermal movement. Stationary blading is carried

bythe innershell. The stationary blade rowsegments of the LP stages are bolted

tothe outer shell of the inner casing. The complete inner casing is supported.

The design of low pressure cylinders has changed a lot in recent years.

Before the advent of the 500 MW machines, condensers were invariably situated

beneath the low pressure turbine and the condenser tubes were at right angles to

the.axis of the machine. With the development of the 500 MW machines several

variations of the above turbine / condenser arrangement have been adopted (Fig.

3.8) shows one such variation with condensers mounted on each side of the I P.

casings. These are called pannier condensers.

PAGE 26

FIG. 3.8 L.P. CYLINDER WITH PANNIER CONDENSERS

Atmospheric relief valve 1 diaphragms (Fig. 3.9) In the event of failure of low vacuum trips the pressure in LP turbine exhaust rises

to an excessively high level. This results in increase of exhaust hood temperature.

The radiation heat due to this can result in loosening of LP rotor blades thus



increasing vibration levels. To protect against excessive internal pressure,

atmospheric relief valves / rupture diaphragms are provided in the exhaust hoods.

Each valve assembly has approximately 1 mm thick gasket clamped between

valve seat and valve disc. In case of high pr. (say 1.2 bar abs.) the valve disc tries

to lift and thereby ruptures the gasket ring, thus allowing the steam to exhaust into

the atmosphere.

3.2 FIXED POINTS, CASING AND ROTOR EXPANSION (500 MW KWU UNIT) In designing the supports for the turbine on the foundation, provisions are kept for

the expansion and contractiion of the machine during thermal cycling.

The fixed points of the turbine easing on the foundation are as follows:

i) The bearing pedestals between the IP and LP turbines. From this point the

IP and HP easing expand towards the front bearing housing of the HP

turbines

ii) LP - Generator bearing housing.

iii) The middle portion of each longitudinal girders of LP turbine. From these

points longitudinal girders expands in both the directions (TS & GS).

PAGE 27

Casing Expansion



The front bearing housings of the HP and IP turbines can slide on their base plates in an

axial direction. Any lateral movement perpendicular to the machine axis is prevented by

fitted keys. The bearing pedestals are connected to the HP & ]P turbine casings by

guides which ensurethattheturbine casing maintain their central position while at the

same time allowing axial movement. Thus the origin of the cumulative expansion of the

casings is at the front bearing pedestals of the LP turbine.

The outer casing of the LP turbine is located axially by fitted keys at the middle of

their longitudinal beam members. Free lateral expansion is allowed. The centre guides f

or these longitudinal beams are recessed in the foundation. There is no restriction

on'axial movement of the casings. At the front and rear supports of the longitudinal beam

members the casing is free to expand horizontally in any direction.

Hence, when there is a temperature rise, the outer casing of the LP turbine expands

from its fixed points towards the generator. Differences in expansion between the outer

casing and the fixed bearing pedestals to which the housings for the shaft glands are

attached are taken by shaft seal compensators.

Rotor Expansion The thrust bearing is incorporated in the front bearing pedestals of the 1 P turbine. Since

this bearing pedestal is free to slide on the baseplate the shafting system moves with it.

Seen from this point both the rotor and casing of the HP turbine expand towards the front

bearing of the HP turbine.

PAGE 28

3.3 MATERIAL FOR CASING AND DIAPHRAGMS

H.P. & I.P. turbine casings and diaphragms are normally Cr, Mo, V creep

resistance steel castings. L.P. casings and diaphragms where the temperature

never exceeds 230'C (e.g. L.P. cylinder on non-reheat machines) are sometimes

made of cast iron. On large reheat turbines, however the temperature of steam

entering the L..P. cylinder may be more than 230'C and because of this and the

large over all dimensions of L.P. cylinders, these are usually fabricated from

carbon steel castings or M.S. Plates, This construction also provides greater

protection in the event of blading failures and speeds manufactures.

3.4 STEAM CHESTS AND STRAINERS

3.4.1 HP Steam Chest



Steam is admitted to a turbine from the steam main pipework via a steam chest,

which normally houses a steam strainer, an emergency stop valve and one or more

governing valves. In some designs a combined emerency stop and control valve is

provided. There may be one or two steam chests provided, depending on the size

of the turbine. The chest consists of a steel casting or, in the case of very high

steam conditions, of a solid forging. Where nozzle govering is used, it is often

convenient to include the governing valve in the high pressure cylinder castings.

Usually the steam chest is firmly anchored to the supporting steel work and

connected to the turbine by loop pipes which are long and flexible enough to allow

the turbine to expand freely. Some manufacturers prefer the chest to move with

the turbine when ft expands; in such cases the movement has to be taken up by the

flexibility of the high pressure steam mains. This has the disadvantage that thrusts

from the steam mains may be transmitted to the turbine. It does, however, permit

shorter loop pipes so that during load rejections. less steam is trapped by the

governing valves and the tendancy to overspeed is lessened.

PAGE 29

A better way of controlling over speed, which becomes even more important as

the size of the machine and steam conditions rise, is to mount single governing chests



on the cylinder. The length of the loop pipes is then of less,consequence, but the

relay gear becomes slightly more difficult to accommodate satisfactorily.

To prevent small lumps of dirt being carried into the turbine by incoming

steam, and causing damage tq. the blading, the steam passes through acylindricai

strainer (which surrounds the stop valve) having holes i about 5 mm diameter. The

danger is particularly prevalent on newplant where pieces of weld metal@and oth(

debris may be swept through the steam mains. When first commissioned, therefore,

extra fine strainer having holes of about 1.5 mm diameterr are fitted. Even particiesd

of this size can cause serious blad damage and thorough blowing out of all pipework

is essential when commissioning new plant.

One of a pair of steam chests fitted to a 500 MWturbine is shown in (Fig. 3.1 0

and 3.1 1). One steam che is fitted to each side of the high pressure cylinder, each

chest containing two emergency stop valves controlling the supply of steam to the

high pressure turbine.



3.4.2 IP Steam Chest

On reheat units it is necessary to provide valve gear to control the flow of steam

from the-reheater into the intermediate pressure cylinder. This valve gear is

housed in a reheat steam chest situated adjacent to the intermediate pressure

cylinder. Large units have two of these steam chests, one mounted on each side of

the, machine.

The reheat steam chest houses an intercept valve which, under certain

conditions, controls the flow of steam, and an emergency stop valve which shuts

off the steam in the event of the turbine valve gear tripping. A diagram of the

reheat steam chest of a 500 MW machine is shown in (Fig.3.12).



PAGE 30

Like some H.P. steam chest the I.P steam chests may also have a combined stop and

control valve.

A reheat strainer is fitted immediately before the steam chest to protect the valve & IP

bladign from damage by debris. A typical strainer is shown in (Fig. 3.13).

PAGE 31

3.5 STEAM VALVES

A turbine is equipped with one or more emergency stop valves, in order to cut off

the steam supply dun'rn pe~of shut down and to provide prompt interruption of the



steam flow in emergency. In addition govemirl valves are used to provide

accurate control of steam flow entering the turbine. Reheat turbines require add~

ermrgency and interceptor valves . in the return path from the reheater and dual

pressure turbines require two @ of emergency and governing valves. (Fig.3.14)

shows some basic schematic designs ol valves in modem use.

a) Shows a “ double--beat" valve having two seatings, the object being to balance the

forces due to steam pressure. It is suitable for most pressures, but not for high

temperatures as differential expansion between the valve and cage would cause

one or other sealing to owrapm.

b) Shows another double-beat valve of the hollow type in which the steam from one

sealing is led through the centre of the valve. The thinner walls promote even

heating and lesser differential expansion.

c) Shows a modern sphedcal valve used for control linghigh temperature steam.

Beinga' sing @beatvalve with one seating, the pressure forces are not balanced and

a large operating force is required.

d) Shows a similar valve fitted with an internal pilot valve which, by opening first,

equalises the pressures and provides initial fine control.

e) Shows a cylindrical valve in which steam pressure is prevented from acting on the

back of the valve by fine annular clearance.



f) Shows a flap valve, used for reheat emergency valves, where the steam pressures

are moderate and the specific volumes (and hence the valve diameters) are large.

g) Shows a governing valve of the *mushroom' type, with a profiled skirt to give a

more linear area 1 lift relationship.

Other types of valves, such as piston and grid valves are used in pass-out turbines.

The diameters of valves opening are generally calculated to dive maximum steam

veloc~ of about 60 m / sec for. ermrgency valves, and about 120 m 1 sec for

governing valves.

The seating upon which any such valves closes is invariably part of a

removable sleeve which is replaceable when worn. -rhe mating annular faces of

valves and their seats are nitrided or faced with Stelide to resist wear. Such wear

is due more to erosion by the steam than to mechanical impact and is particulars

lipme to taker place when the valve is cracked open and a jet of steam is propelled

at high velocity through the n arrow port opening by the large pressure differential,

impact damage can occur as a result of frequent test ck)sures, and cushioning

devices or slow motion testing may be adopted to avoid this.

PAGE 32

3.5.1 Valves in the Steam Chests

Depending on the size and design the number of types of valves in steam chests

will vary. For example the valves where as the HP turbine of same rating from

BHEL / KWU design has only two combined ESV and control valves.

Similarly the type and number of valves in IP steam chest will also vary

according to designers choice. All these valves are quite massive in size and are

operated by lever and hydraulic relays which inturn are guided by the governing

system. (Fig. 3.15 to 3.18) shows the lever arrangement for these valves.



PAGE 33



PAGE 34

3.6 LOOP PIPES

Steam passes from the steam chest to the turbine via loop pipes which are normally

U-shaped to give them sufficient flexibility (ft is important that these loops be

provided with drain rocks for use when starting up). With the use of high pressures,

the pipe walls have to be thick, making the pipes stiff. To achieve the required

flexibility and to avoid the imposition of large forces or bending moments on the

turbine very long loops are required.

Where pipes enter a double shell cylinder, it is preferable that they enter

radially, passing through asliding joint in the outer cylinder; in this way the two

sheli scan expand radially without losing concentricity. The sliding joint usually

contains piston rings made of nimonic alloy or special steel which will retain its

springiness at the prevailing steam temperature. See (Fig. 3.19 (a).



Cross-over pipes between cylinder must also be flexible, as they expand more

than the bearing pedestal and cylinders over which they pass. Pipes with long loops

are used for transmitting very hot steam. Where possible, crossover pipes pass under

or alongside cylinders rather than overhead, to improve cylinder access.

Expansion of LP cross over pipes is taken up by two or more hinge-linked

bellows which allow bending but no axial movement (Fig. 3.19 (b) in this way the

pressure force in the pipe is transmitted through the links, thus protecting the

convolutions from the tenency to open out. Afternatively, straight linked bellows may

be used in pairs, as shown in (Fig. 3.19 (c).,

PAGE 35 3.6 ROTORS 3.7.1 There are two types of turbine rotor used in large turbines which have impulse type

blading : a) The built up rotor also called Disc Rotor consisting of a forged steel shaft on

which separate forged steel discs are shrunk and keyed. (Fig. 3.20).

b) The integral rotor in which the wheels and shaft are formed from one solid

forging. (Fig. 3.21).

The built up rotor is made up of a number of separately forged discs or

wheels and the hubs of these wheels are shrunk and keyed on to the central shaft.

The outer rims of the wheels have suitable grooves machined to allow for fixing

the blades. The shaft is sometimes stepped so that the wheel hubs can be threaded



along to their correct positions. Suitable clearances are left between the hubs to

allow for expansion axially along the line of the shaft.

PAGE 36

Integral rotors as said before have discs and shaft machined from one solid

forging, the whole rotor being one complete' icce of metal. This results in a rigid

construction and troubles due to lobse wheels of the shrunk on type are eliminated.

Grooves are machined in the wheel rims to take the necessary blading. These are

also called solid forged rotors.

FIG. 3.21 INTEGRAL ROTOR

The built-up rotor tends to be the cheaper of the two since the discs and

shaft are relatively easy to forge and inspect for flaws; also, the machining of these

components can be carried out concurrently. On the other hand, integral rotors are

expensive and difficult to forge and there is a high incidence of rejects; there is

also a large amount of machinery time and waste material involved.

In spite of the expenses involved, the advantages of integral rotors are such

that they are invariably used for the high pressure rotors on high temperature

plant; on reheatmachines in particular they are often used for intermediate pressure

and low pressure rotors as well. This is because of the difficulty of ensuring that

the shrunk-on discs on intermediate and low pressure rotors cannot become loose,

particularly at the high temperature end during start up when the shafts may be

relatively cool and the discs are hot. Another source of trouble under conditions



of high temperature and stress is the phenomenon of creep which could also cause

the shrink-fft to disappear after a large number of running hours.

With regard to low pressure rotors, the main problem is one of centrifugal

stress, the last stage being the most heavily stressed part of the turbine. The last

row wheels on the standard 500 MW turbine are the largest cap able of operating

at 1 000 rev 1 min; the blades are 900 mm in length and are mounted on the disc

so as to have a mean diameter of 2.5 m, the overall diameter is therefore 3.45 m.

On large turbines using 50 per cent reaction, four types of rotor are used:

a) The hollow drum rotor which promotes even temperature distribution because it is

designed with the same thickness of material as the casing.

(Fig. 3.22) illustrates the construction of the hollow drum type rotor.

b) The solid drum rotor suitable for cylinders where there are lower temperatures but

large diameters, as in intermediate pressure cylinders without reheat.

c) The built up rotor previously described.

d) Welded Rotors which are built up. From a number of discs and two shaft ends.

These are joined together by welding at the circumferences and because, there are

no central holes in the discs the whole structure has considerable strength. Small

holes are drilled in the discs to allow steam to enter inside the rotor body to give

uniform heating when coming on load. Grooves are machined in the discs to carry

the blades and (Fig. 3.23) shows this type of rotor construction.



PAGE 37



PAGE 38

3.7.2 Balancing

When assembled complete with blades the rotor is balanced both statistically and

dynamically.

In the case of built-up rotors, each bladed disc is balanced individually

before assembly. The aim of balancing is to reduce vibration to a tolerable level,

usually accepted to be about 35μm at the bearing pedestal of a 3000 rev / min.

machine.

A stationary shaft supported between bearings has a natural frequency of

vibration depending upon its diameter and the distance between the bearings. If its

speed of rotation corresponds to its natural vibration frequency, the residual out –

of-balance forces can build up to a dangerous extent. This speed is known as the



Critical Speed. Sometimes it is above the running speed and sometimes it is

below. If above, the shaft is said to be stiff, if below, the shaft is said to be flexible

and the critical speed should be passed as quickly as possible when running the

turbine up to speed.

The critical speed should not be within 20 per cent of the running speed.

PAGE 39 3.7.3 Materials

High and intermediate pressure rotors are usually made from chromium-moly

bdenum-vanadium steel ferritic materiao which is suitableforwheeicase steam

temperatures up to 540'C. Austenitic materials are no favoured because of the cost

of their manufacture and their high coefficient of expansion.

Low pressure rotors are made of 3 per cent chromium-molybdenum steel or

2 114 per cent nickd chromium-molybdenum steel. The first of these is not used

in modern designs because of the tendenq towards scuffing in the bearings during

the early life of the rotor.

Low pressure discs are made of 3 per cent chromium-molybdenum-

vanadium steel, or more recentn 0 3 112 per cent nickel-chromium-molybdenum-

vanadium steel.

3.8 BLADES

These are most important (and co" too) components of the turbine as these are

respcinsible for the ma:i@ function of the turbine, i.ib. conveting heat energy to

mechanical energy.

A blade has three main parts:

- AEROFOIL - It is the working part of the blade

- ROOT - It is the portion of the blade which is fixed with the

rotor or casing.

- SHROUD - It can be rivetted to the main blade or can be integrally

machined with the blade

(Note: Blades maybe without shroud also).



PAGE 40

3.8.1 Type of Blades

Most modern turbines use reaction type blading throughout the machine.

Some designs have impulse in the H.P. and I.P.cylinders and reaction in the L.P.

cylinder.

But use of impulse or reaction cannot always be dearly defined because

both principles may be combined in the same blade. For example large L. P.

blades are generally of twisted and tapered design (see fig. 3.24). These blades

produce varying conditions of impulse or reaction between root and tip and are

called vortex blades. The object of this design is to prevent uneven steam flow

caused by centrifugal forces forcing the steam towards blade tips. This is done

by changing the throat opening from root to tip. A 915 mrn (36m) blade with

zero reaction at the root has approximately 70 per cent reaction at the tip. Also



the inlet angle of the blade afters along its length giving smooth and efficient

steam entry.

3.8.2 Impulse Type Moving Blades

lmpluse type moving blades (for H.P.. Turbine) are machined from solid

bar and the roots and spa@rsformed with the blade (Fig. 3.25). Tangs are left at

the tips of the blades so that when fitted in position in the wheel, shrouding can be

attached.

The shrouding is made up from sections of metal strip punched with holes

to correspond with the tangs. As there is no pressure drop across the moving

blade, the seating arrangements are not of such great importance, as in the reaction

type. The shrouding on the impluse blading helps to guide the steam through the

moving blades, allowing larger radial clearance, as well as strengthening the

assembly.

PAGE 41

3.8.3 Impulse Type Fixed Blading (Fig. 3.26).

The fixed blading in an ampluse turbine takes the form of nozzles mounted in

diaphragms. The diaphragm is made in two halves, one half being fixed to the

upper half of the cylinder casings by means of keys so that when expansion

occurs fouling of the shaft seals is avoided. Special carrier rinfs are generally used

to support the diaphragms in H.P. cylinders.

Because of the steam pressure difference on each side of the diaphragm,

seals are provided at the bore where the shaft passes through the diaphragm, to

prevent steam leakage along the shaft.



FIG. 3.26 IMPULSE TYPE FIXED BLADING

In reaction type blading pressure drop occurs across both the fixed and

moving blades. So, very effective seal between fixed and moving blading is

essential to prevent steam leakage which would make the turbine inefficient. The

leakage of steam controlled by axial clearance is shown in (Fig. 3.27). This type

of sealing is called end tightening. Following is the details of Reaction type

blading of the H.P. Turbine of 210 Mw Set (KWU / BHEL).

PAGE 42



FIG. 3.27 REACTION TYPE BLADING-SHOWING END TIGHTENING

Moving and Stationary Blades

The HP turbine blading consists of several drum strages. All stages are reaction

stages with 50 per cent. The stationary and moving blades of the front stages (Fig.

3.28) are provided with T-roots which also determine the distance between the

blades. Their cover plates are machined intergral with the blades and provide a

continuous shroud after insertion.

PAGE 43

The moving stationary blades are inserted into appropriately shaped grooves

closed casing (1) and are bottom caulked with caulking material (9). The insertion

slot in the shaft (8) i a locking blade which is fixed either by taper pins or grub

screws. Special end blades which lock with t horizontal joint are used at the

horizontal joints of the inner casing. Graub screws which are inserted from t joint

into the material secure the stationary blades in the grooves.

Gap Sealing

Sealing strips (3,7) are caulked into the inner casing (1) and the shaft (8) to reduce

leakage and losses at th blade tips (4,5). Cylindrically machined surfaces on the

blade shrouds are opposite the sealing strips. Thesf surfaces have stepped

diameters in orderto increase the turbulence of the steam and thus the sealing

effec@ Should an operation disturbance cause the sealing strips to come into

contact with opposite surfaces they ar@ rubbed away without any considerable

amount of heat being generated.

3.9 SPECIAL FEATURES OF LP TURBINE BLADING

The blades used in LP Turbine have to be very long to cater for increasing specific

volume of steam at the lower pressures. In the final stages of LP turbine the prime

importance is of ensuring uniform steam flo@ around the circumference. To

achieve this twisted and ta@ered blades are generally used (function of whic@ area

already explained) to dampen vibration, the long blades may be laced together (Fig.

3.29). But as facing holes are a source of weakness and wires upset the steam path,

the trend now a days is to avoid lacing b@ designing blades whose natural

frequency is well clear of the rate of impingement at running speed.



In the LP turbine stages the steam becomes wet. Maximum wetness normally

tolerated is about 12 pe@ cent. In wet region some steam condenses to tiny

droplets of water and water quantity increases as stean continues its expansion.

Water droplets, being heavier, flying outwards by centrifugal force. Part of it is

conducted away by steam extraction with water catch channels and part collected in

special draining grooves built into the turbine casing (Fig.3.30).



PAGE 44

Condensing droplets impinge on the moving blades and cause erosion on

the leading edges of the (mainly) last few stages. So, generally, physical damage



to moving blades of last two stages are minimized by fitting “Stellite” shields on

the inlet edges (stellite is the proprietary name for a series of extremely hard alloys

of cobalt, chromium, tungsten and carbon). (Fig. 3.30) also shows the stellite

protection provided to the last two blades of 210 MW BHEL / LMW Turbine.

In other design (Fig. 3.31 of 210 MW BHEL / KWU set) the fixed blades

of last two stages are hollow. Slits are provided in the blade surface of last

stationary row through which any water passing over the blade can up of any

droplets which may still remain. The leading edges of the final stage rotors are

flame hardened to give protection against erosion.

PAGE 45

3.9.1 Baumann Exhaust 3.1

A special type of exhaust blading, invented by Dr. Baumann, is used by some

manufacturers to achieve 3.1 greater exhaust area wmmtd undue lengthening of

the last row blades. I ncrerased exhaust area is favoured to reduce 'leaving loss'.

Also known as multi exhaust, it uses two-tier blades for the penultimate stage.

(Fig. 3.32) shows the fast but one stage of a turbine fitted with a Bauniann

exhaust. As the steam passes through the turbine to the exhaust, the blading is

conventional until it reaches the fixed blade A. This blade has a circular di~on ring

B which spins the steam into two paths. The outer path of steam is expanded to

back pressure and passed to the condenser via the moving blades C. Meanwhile

the other steam is discharged from A at a pressure higher than the exhaust

pressure. It p through the penultimate moving row of blades, after which R passes

to a further rcww of fixed blades D contained within a shroud E. Here the pressure

is reduced to that of the exhaust before. being discharged to the condenser via the

last row of blades F.



PAGE 46

3.10 METHODS OF FITTING BLADES

3.10.1 Root Fixings

Many types of root fixing shapes exist for turbine blading to suit both the

conditions u nder which the blade must operate and the preference of the particular

designer concerned. In general there are the types which either fit

intheirappropriate groove or straddle ft, whilst otherdesigns are fixed by rivets

through the blade root. Some examples of these blade root fixings are shown in

(Fig. 3.33).



In the case of reaction type turbines the H.P. blading is built up in packets

of up to ten blades and held in the rotorgroovewith serrated looking stripsas shown

earlier in (Figure 3.27). A gap is left in the rotorgroove to allow the last serrated

locking strip to be inserted. This gap is then closed with a plate fixed by screws to

the rotor body. In these turbines, the last L.P. rotor blades are subjected to a

centrifugal force of about 250 Tons / blade. Here fir tree fixing has proved most

effective. This is also shown at (Fig. 3.31).

For some impulse type blading a gap or "gate" is left in the rotor groove

and the blades.,are fed in and located by the particular shape of serrations used for

the blade root and groove. The last blade, which has a plain root with no

serrations, fits over the "gate" and is secured by riveting.

In other methods of fitting a soft metal distance piece or wedge is driven in

the groove to locate the last blade.



Where the blades are all attached to the rotor wheels by riveting or are of

the side entry type any blade can be last one and each must be separately secured

by riveting or with locking plates.

PAGE 47

3.11 SHAFT TURNING (BARRING) GEAR

Turning gear is provided to rotate turbine shafts slowly during the pre-run up

operation and after shut down to prevent uneven heating or cooling of the shafts.

The uneven heating or cooling would lead to bending and misalignment of shafts

with possible fouling of stationary and moving parts.

Use of turning gear during starting eliminates the necessity of admitting

suddenly a large flow of steam to rotate the turbine from the rest.

The turning gear speed is chosen to ensure satisfactory lubrication of the

bearing and, at the same time, provide some circulation of air within the casing

(particularly at the low pressure end) after shut down. The speed of turning gear

varies considerably from one design to another. For example while BHEL 1

LMW 210 MW turbine is rotated by the turning gear atthe speed of 3.4 rpm, in

500 MW KWU turbines, the T/ G rotates the turbineshaft at 270 R.P.M. / or 240

R.P.M. depending on whether the condenser is under vacuum or not.

The turbine must remain on turning gear units metal temperature has

dropped below 150"c with normall cooling, this will take approximately 72

hours.

Before putting the turbine on turning gear a few conditions like-adequate

bearing oil pressure, jacking oil pump running etc. must be satisfied.

JACKING OIL PUMPS (JOPS) are positive dispi - acement pumps that

provide high pressure (1 20 bar for KWU turbines) supply of oil under strategic

journals @~turbo generator and the oil lifts the shaft slightly, This ensures that

there is no metal contact between a journal and the bearing. This greatly reduces

the static friction and bearing wear, also the starting torque headed by the turning

gear drive. The JOP can be stopped after the lubricating oil film is established

between the shaft and bearings.

On early turbogenerators, turning of TG was done by hand withthe help of

long barfitted with rachetworm and pinion mechanism. (Fig. 3.34). This gave the



name "barring gear" or "barring" to this operation. Now, the driving force is

provided by either electric motors or hydraulic pressure.

Hand barring gear is used, nowadays, in emergency, when T.G. motor is

non-operational, and for

PAGE 48

maintenance purposes, to rotate the turbine shaft manually. An auxiliary source of

power from U.P.S. (Uninterrupted Power Supply) or Diesel Generating set is also

provided in some cases for reliability of T /G operation.

Fig.3.35 shows the functional arrangement of a turning gear.

In BHEL 1 21 0 MW LMW turbine, the T./ G is mounted on LP rotor rear

coupling. It consists of a worm,,



worm wheel spur gear and pinion, spiral shaft and sliding shaft with lever. The

system comes into operation when the shaft comes to stand still. When T / G is

engaged, the turbine shaft rotates at 3.4 R. P. M.

In KWU turbines, the turning gear is hydraulic. It is engaged when shaft

speed comes down to 545 R.P.M The T 1 G rotates the shaft at 120 RPM or 80

R.P.M. depending on whether the condenser is under vacuum or not. The T / G

assembly is located in the front bearing pedestal of LP cylinder and consists of two

rows of moving blades mounted on coupling flange of I.P. rotor, an inlet nozzle

box with stationary nozzles and .,guideblades (Fig. 3.36). The TG shaft system is

rotated by the double row wheel which is driven by prerssurised oil supplied by

auxiliary oil pump. After passing through the blading the oil drains to the bearing

pedestal and combines with the bearing lube oil returning to the iube oil tank.

In addition, the system is equipped with facility for manual barring in the even

of failure of hydraulic turning gear.

PAGE 49

3.12 COUPLING

3.12.1 Introduction

The need for couplings arises from the limiting length of shaft which it tis possible

to forge in one piece and from the frequent need to use different materials for the

various rotors, in view of the various conditions of temperature and stress.



Couplings are essentially devices fir transmitting torque, but they may also have to

allow relative angular misalignment: transmit axial thrust, and ensure axial

location or allow relative axial movement. They may be classified as flexible,

semi-flexible or rigid.

Type of coupling used may vary from manufacturer to manufacturer. For

example, the BHEL / LMW 210 MW units employs a rigid coupling to connect

HP and IP Turbine and a semi flexible one for connecting IP and LP turbine;

whereas, both the couplings of 210 MW BHEL /KWU set are of rigid type.

Following are the brief descriptions of basic three types of couplings;

PAGE 50

3.12.2 Flexible Coupling

Flexible couplings are capable of absorbing small amounts of angular misalignment

as well as axial movement. Double flexible couplings can also accommodate

eccentricity. Semi-flexible couplings will allow angular bending only.

(Fig. 3.37) shows some designs in common use. The claw coupling, which

may be single or double, is robust and slides easily when transmitting light load; on

heavy load, however, friction causes ft to become axially rigid. The Bibby

coupling is satisfactory up to medium sizes and provides, in addition to the other

features, torsional resilience' The mutti tooth coupling transmits torque by internal

and external gear teeth of involute form, which are curved to accommodate angular

misalignment,

All these couplings require continuous lubrication, normally obtained from

a jet of oil feeding into an annual recess, from which k is led centrifugally to the

coupling teeth through drilled passage_ways.



PAGE 51

3.12.3 Semi flexible Coupling

The semi-flexible type of coupling requires no lubrication and is normally

interposed between the turbine and generator. It consists of a hollow piece having

one or more convolutions (Fig. 3.38. (a)

3.12.4 Rigid Couplings

On large turbines the high torque to be transmitted renders the use of flexible

couplings impracticable. Consequently rigid couplings are employed between the

turbine cylinders so that the turbine shaft behaves as one continuous rotor A spigot

locates the two half-couplings and numbered fitted bolts join the flanges (Fig. 3.38

(b).



PAGE 52

GLANDS AND GLAND SEALING SYSTEMS

4.1 GLANDS

Glands are used on turbine to prevent or reduce the leakage of steam or air

between rotating and stationary components which have a pressure difference

across them; this applies particularly where the turbine shaft passes through the

cylinder. If the cylinder pressure is higher than atmospheric pressure there will be

a general steam leakage outwards; d the cylinder is below atmospheric pressure

there Mil be a leakage of air inwards, and some sort of sealing system must be

used to prevent the air from entering the cylinder and the condenser.

4.1.1 Water-sealed glands

Some turbine designs incorporate a shaft gland which depends on a water seal to

prevent steam or air leakage. A typical seal arrangement (Fig. 4.1) consists of a

shaft - mounted impeller with a series of vanes or pockets machined on both faces.

The impeller is contained within an annular chamber, and, when water is admitted

to the chamber, the impeller vanes force the water to rotate, at a speed

approximately equal to the impeller speed. The seal is relatively inefficient at low

speeds and air-sealed auxiliary labyrinth glands must be used, in conjunction with

high capacity air pumps, to raise vacuum when starting. Water is usually injected

into the seal at approximately hag of the full operating speed.



FIG. 4.1 A TYPICAL WATER-SEALED GLAN FIG. 4.2 (A, B & C) LABYRINTH GLAND DESIGN

PAGE 53

The side clearances between the impeller and seal chamber must be fairly

small, and so the use of this seal is restricted to positions on a turbine where the

axial differential expansions are within the effective limits of impeller and seal

chamber clearance. When this type of seal is used on a high pressure turbine, the

seal cannot absorb the full differential pressure so air-sealed labyrinth glands are

used to breakthe pressure down to a figure which the water seal can handle.

Since a water seal absorbs and generates heat, the water contained in the

annular chamber of the water - sealed gland is continuously evaporated; the water

losses are made up from a header tank.

4.1.2 Labyrinth glands

In modern turbines the labyrinth gland are used because it can withstand high

pressures and temperatures and yet requires little maintenance.

The labyrinth gland provides a series of very fine annualar clearances, in

the gap between the cylin@, wall and the shaft. The steam isthrottled through this

gap and its pressure reduced step by step. In expandii@,, through each clearance,

the steam develops kinetic energy at the expense of its pressure energy; idealism



the kinetic energy is converted by turbulence into heat with no recovery of

pressure energy. In this way, t@ @e pressure is progressively broken down as the

steam is throttled at successive restrictions. By keeping t,@,,, clearance area

sufficiently small, the quantity of energy lost may be kept low, and as increase in

turbine outr@l occur the gland leakage loss becomes proportionately less.

To reduce the clearance area, glands are made with a diameter as small as

possible, and clearances . @s fine as possible. The diameter is limited by

considerations of shaft strength and radial clearance, by t clearance within the

bearing, and by the possibility of shaft distortion.

Glands must allow for axial expansion of the shaft and casing to take place

without causing a rub. On t' other hand, if a rub does take place because of shaft

vibration, ft is desirable that the heat generated minimized to prevent serious

frictional heating of the shaft and possible distortion. A typical modern giants,.

comprises stationary fins on spring-loaded sectors, white the shaft is either smooth

or castellated. If a should occur, the sectors receive the generated heat and can be

replaced readily if they are damaged.

Designs of labyrinth gland at present in use are shown in (Fig. 4.2).

In (Fig. 4.2(a) the clearances are staggered to ensurethat no kinetic energy is

carried overfrom one genius to the next. The stationary fins are axial, so that if a

rub occurs, the heat causes them to expand relative their fixing, and they move out

to increase the clearance.

(Fig. 4.2 (b) shows a resilient gland, the stationary part being divided into

sectors, each of which is sprir@,' loaded in an annular groove. If a rub occurs, the

sectorwould'give', andthe low contac that only a little heat is generated.

The gland shown in (Fig. 4.2 (c) is of the vernier type, the fins being much

finer than in the previo,,@,designs. By making the pitch of the fins on one side 1

0 per cent greater than the pitch of the mating fins, oi-@ one fin in nine or ten will

be opposite another fin. If a rub occu rs, then only exactly opposing fins make

conta so the amount of heat generated is small.

Each of these glands can accommodate a certain amount of axial movement of

the rotor without dama or loss of effectiveness.



The arrangement of glands and gland sealing system vary between different

makes and sizes. Though purpose of the system is same. Following is the

description of the gland sealing system of a 21 0 MW (LMW 1 BHEL) turbine.

Gland sealing system (21 0 MW LMW 1 BHEL Turbine) (Fig.4.4)

The 210 MW turbine has got 6 sets of gland seal one for each side of HP and LP.

Each set of gland seal consists of no. of sectionalised gland. HP turbine front

gland is sectionalised in to five section. HP rear and IP front into four sections, IP

rear into three sections and LP glands into two sections. Each gland sealing

consists of number of sealing rings divided into segment, each segment is backed

by two flat springs. The no. of sealing rings depended on pressur e against which

it is working. The sealing rings are housed in grooves machined in gland bodies

which in turn are housed in the turbine casing or bolted to the casing at ends.

The glands are sealed by steam. Penultimate section of each gland is

supplied with steam from gland

PAGE 54

FIG.4.3 HIGH PRESSURE CYLINDER LABYRINTH GLAND (PARCON TURBINE)

Steam header maintained at a pressure of 1.03Kg/cmlabsto 1.05Kg/CM2absand

between 130'>cto 1500c. The header receives steam from de.3erator steam space



through a pressure control valve. Provision is also there to supply the steam from

TAS (Turbine Aux steam) system during non availability of deaerator steam.

Steam is suppliedto last but one seal of each gland at a pressure of 0.1 Oto

0.20 Kg/ CM2 and temperature 130'cto 1 50'c from sealing steam header. Leak

off steam from the turbine istapped off from different section in the gland seal and

is either cooled in the gland steam cooler or fed to lower pressure stages of

turbine. The leakage from the last stage i.e. air side leakage is cooled in gland

cooler No. 1. The cooler is maintained under vacuum with help of a special steam

ejector provided inside the Gland Cooler No.l.

Steam from fourth sealing chamber (from air side) of HP front, rear and IP

front is connected with turbine 4th Extraction before the NRV. Steam from the

third chamber of HP front, rear and I P front rear is connected with the Gland

steam cooler No,2 for regenerative feed heating cycle. The third gland chamber

and gland cooler No.2 are always maintained at condenser vacuum as the

condensate drain side of gland cooler No. 2 is connected with the condenser. The

lead off steamfrom thefirstchamber of the HP front gland is connected to HP

turbine last stage. Steam / air mixture from the spindle seals of ESVS. IVs, and

control valves of HP and IP is exhausted into the gland steam cooler No.l.

In addition to the above mentioned gland steam supply system, an another

source of supplying gland air from the live steam is there.

All the gland seals of turbine are normally fed with steam from deaerator at

controlled pressure. Following a turbine trip-out, this gland steam shall be sucked

into the cylinders which would be under vacuum at that condition. Sincethe

temperature of gland steam fed from deaerator is much lowerthanthe

steamtemperature at ini6ts to HPT and IPT, differential expansion of HPT and IPT

are likely to increase at a faster rate under such conditions. In view of the above a

provision has been made to inject main steam to the front gland seals of HPT and

IPT. This system is called Rotor heating system.



PAGE 55

PAGE 56

CONDENSATE SYSTEM

A typical condensate system consists of the following:

i) Condenser (including hot-well)

ii) Condensate pumps

iii) Air Extraction System

iv) Gland coolers and L.P. heaters

v) Deaerator

5.1 CONDENSER

The functions of condenser are:

i) To provide lowest economic heat rejection temperature for the steam. Thus

saving on steam required per unit of electricity.

ii) To convert exhaust steam to water for reuse thus saving on feed water

requirement.

iii) Deaeration of make-up water introduced in the condenser.

iv) To form a convenient point for introducing make up water.



Type of Condenser

Condenser is basically a heat exchanger and hence can be of two types:

i) Direct contact

ii) Surface contact

5.1.1 Direct Contact Type (Jet Condenser)

In this type, condensation of steam takes place by directly mixing exhaust steam

and cooling water. Requirement of cooling water is much less here compared to

surface type. But cooling water qualityshould be equal to condensate quality (Fig.

5.1).

5.1.2 Surface Condenser

This type is generally used for modern steam turbine installations. Condensation

of exhaust steam takes place on the outer surface of the tubes which are cooled by

water flowing inside them (Fig.5.2).

The condenser essentially consists of a shell which encloses the steam

space. Tubes carrying cooling water pass through the steam space. The tubes are

supplied cooling water from inlet water box on one side and discharged, after

taking away heat from the steam, to the outlet water box on the other side.

Instead of one inlet and one outietwater boxes, there may be two or more

pair of separate inlet-outietwater boxes, each supplying cooling water to a separate

bundle of tubes. This enables cleaning and maintenance of part of the tubes while

turbine can be kept running on a reduced load.

5.1.3 Description of Condenser for 210 MW (BHEL) Turbines (Fig. 5.3)

The condenser group consists of two condensers, each connected with exhaust part

of low pressure casing. These two condensers have been interconnected by a by-

pass branch pipe. The condenser has been designed to create vacuum at the

exhaust of steam turbine and to provide pure condensate for reusing as feed

waterforthe boilers. The tube layout of condenser has been arranged to ensure

efficient heat transfer from steam to cooling water passing through the tubes, and

at the same time the resistance to flow of steam has been reduced to the barest

minimum.



350% capacity condensate pumping sets are installed for pumping the

condensate from condenser to the deaerator through low pressure heaters. Two

pumps are for normal operation and one works as stand by pump.

PAGE 57

PAGE 58

Constructional feature

Each condenser has been sub-divided into upper and lower parts. Front water box,

shell and rear water box constitute the lowerpart. Two end tube plates and six

support plates are located inside the lower body of the condenser.

Front water boxes have been divided into two parts to make the condenser

two pass design. End covers of water boxes are kept detachable for facilitating

repairs and replacement of tubes. Manholes have been provided for routine

maintenance and visual inspection along with venting and draining arrangement

for individual water boxes. Condensertubes are secured to the end tube plates by



expanding and flanging of tube ends which provides very good sealing

arrangement against penetration of circulating water into the steam space. The

tubes have been so arranged that there is equal distribution of steam on the tube

nest with minimum resistance to steam flow. Non-condensable gases are

continuously sucked with the help of steam ejectors.

With a view to allow relative expansion between tubes and the body of the

lower part, lens type compensator has been provided in the body itself at the rear

water box end. This arrangement prevents deformation of the body and damage to

connections between tubes and end plates.

FIG. 5.3 CONDENSER

Upper part of condenser has been designed to allow smooth flow of steam

over tube nest. It consists of mild steal flat walls, strengthened from inside by

gratings of longitudinal and transverse rods and from outside by channels. These

rigid bars help the condenser to ratain its shape against atmospheric pressure.

Two sections of low pressure heater No, l have also been located inside the

upper parts of condenser. In orderto allow expansion along the height, the

condenser is supported on springs specially designed to take up load (Fig,5.9).

The weight of the condenser and its tubes is taken by the springs and

through them by the condenser foundation. The weight of circulating water and



the condensate alon~h the thrust of springs during expansion is transfered to

turbine foundation.

PAGE 59

Special care has been taken for removal of condensate formed as a result of

condensation of steam. Baffle p"es have been provided to guide the steam flow on

the tube nest and for collecting the condensate h~ from upper rows of tubes and

directing it towards the intermediate support plates for flowing down in na~ layers,

leaving the passage free from steam flow.

Asteam throw off device has been incorporated in each condenser for

dumping the steam into the condenser during start up and sudden load throw off

from the set.

5.1.4 Materials for Condenser Tubes

Selection of tube material depends mainly on the quality of cooling water and the

cost. Copper bearing alloys are prefered as copper has very high beat transfer

coefficient. But as copper has very little mechanical strength; d has to be

reinforced by alloying with other metals. Copper alloys are basically of three

categories. Br (h) C-upro-nickel and (iii) Bronzes.

Stainless steel tubes has also been used and has good corrosion resistance

though heat transfer coeffi~t is quite lower than the copper alloys. Because of

high cost, stainless steel is used only where water is highly corrosive. So~ sea

side power plants are also using Titanium despite high cost, because of highly

corrosive environment.

5.1.5 Tube Packing

The ~hod otattachmentof thetubestothe tube plate is very important. These tubes

being brass will expand more ffim the steel shed when the condenser warms up

under working conditions and allowance must be made for this extra expansion

when the tubes are attached to the tube plate.

The method adopted is to allow the tube to @ids through the tube plate as it

expends. This means that the holes must be bigger ~ the tube and, to prevent

leakage, the gap between tube and tube plate must be sealed. Linen or metallic

paddngs are used for sealing purposes.



The pacidngs are held in place by means of ferrules which slide over the

tube and screw into the tube plate. Lines tape treated with raw linseed oil has

given very satisfactory service, bin matailic packing is often prefered. Metallic

packing gives a firm metal contact from tube to tube plate. This, ft is claimed,

prevents corrosion of the tube ends by ~~c action, which is corrosion due to the

passage of very small electric currents.

When using ferrules, A is important that sufficient space is left at the end to

allow the tube to expend fully. If ferrules are used at both ends of the tube it is

better thal flush ferrules be used at the inlet end to cause least i~rwence wM ~er

flow. (Fig.5.4)

5.1.6 Expand Tubes

Afternatively, the inlet ends of the tubes may be bell mouthed and expanded.

Occasionally tubes are expanded at both ends and in this case expansion must be

allowed for in other ways, for example by a bellows arrangement built into one

end of the condenser.

With double tube places, the tubes are fixed to the tube plates by rolling

into form the required expansion as ~wn in (Fig .5.5).

PAGE 60



5.1.7 Tube Nest Arrangement

In addition to designing the condenser to give a low back pressure while

using as little pumping power as possible, it also necessary to ensure that the

condensate is nor under cooled and tat the pressure drop of the steam path through

the condenser is a small a possible.

In early condensers the tube bundles were tightly packed tigether. As a

result, little steam penetrated to the bottom of the condenser, and most of it

condensed in the upper parts of the condenser. Consequently, as the droplets of

condensate fell through the condenser tube nest and struck more tubes, they were

cooled below the saturation temperature of the steam.

PAGE 61

The first step in improving the tube nest arrangement was to provide considerable

space around the tube bundles and to incorporate wide stearb lanes to allow steam

to,circulate freely (Fig.S.6 ). The steam can penetrate to the bottom, of the

condenser to assist the even heat distribution to the lower passes of the

condenser.'This design allowed only part of the steam to condense in the lower

parts of the condenser with the advantage that those condensate droplets did not

have far to travel; those droplets which fell from the upper parts of the condenser

also had to pass through the warn steam., so helping to reduce under coo ing .



A further improvement was the introduction of condensate deflectors.

These plates or'trays collect the condensate droplets and direct them away from the

lower tubes so that they fall directly into the hotwell again reducing under colling.

In modem condensers particular care is gives to channelling part of the

exhaust steam directly to the space immediately above the hotwelf; the object of

this is to recover as much as possible of its velocity-head energy to achieve the

maximum vapour pressure above the condensate in the hotwell. Some of this

vapour condenses directly on the surface of the condensate in the hotwell. This

belt of higher pressure and higher temperature steam has to be crossed (see

fig.S.7) by the conden sate droplets failing into the hotwell, and so their

temperature is increased.

By improving the steam distribution in the condenser the pressure drop

across the condenser has also been reduced. However,'with the conventional

exhaust arrangement with understung condensers, the necessity to pass the top hag

of the exhaust steam across the horizontal joint imposes a limit on the degree of

pressure recovery that can be obtained economically.

This difficulty has been overcome by raising the condenser from below the

low pressure cylinder to the same level as the cylinder. Under conditions, the

exhaust steam flows directly into the condensing surfaces, with the result that the

length and complexity of the steam path is greatly reduced. (Fig.S.8) and (Fig.3.8)

show two variations of this development. (Fig.5.8) illustrates the principle of the

radial condenser, and (Fig.3.8) the pannier condenser. In case 'of pannier

condensers which are recently gaining popularity, the savings from increase in

efficiency, and savings from a reduction in the basement depth far exceed any

increase in building costs caused by the extra width of the turbine.



FIG. 5.6 IMPROVED TUBE NEST ARRANGEMENT FIG. 5.7 STEAM AND CONDENSATE

FLOWS THROUGH CONDENSER PAGE 62

5.1.8 Condenser leaks

Air leakage into the condenser is one of the main cause of poor vacuum.

5.1.8.1 Location of air leaks

The traditional method of locating air leaks when the turbine is on load is to

pass a lighted taper round the joints which are suspected of having a leak. The

flame of the taper is drawn towards the place where the air is being drawn into

the condenser.



This is a time-consuming technique as the taper has to be passed slowly

over every area where a leak is suspected, and the presense of draughts

cantnake this a very frustrating job.

A quicker way of locating leaks is to spray the suspected area with freon or

other halogen gas. This is then .drawn into the condenser and sucked into the

air extraction equipment. If a lighed blow lamp is placed with its flame above

the air discharge port on the air extraction equipment the normally blue flame

will change to orange when the halogen is emitted.

A more modern development of this method is the use of halogen gas

detectors. These are. inserted into the air discharge line from the air extraction

equipment and a meter registers wh'e"n a halogen gas passes the detector. A

suitable gas (such as, freon) is sprayedround the suspected area until the

detector registers.

The disadvantages of these systems are:

a) The operation needs two men: one, mans praying, and the other watching the

blow lamp or indicator.

b) Time must be allowed to elapse aftere ach spray so that, if there is an

indication, the operator knows which area that has been sprayed contains the

leak.

Off-load leak searches are carried out by filling the condensate system and

stearb space with water to a level below the turbine blades.. Care must be taken to

ensure that the condenser supports have first been set in the correct position to cater

for the extra load in the condenser. (Fig.S.9) shows how the condenser supports are

arranged.

Fluorescene is added to the water, and if any leakage takes place the

fluorescene can be detected by the use of an uftra-violet lamp. Leakage is detected by

this method, not only at the condenser mountings, but also on the low pressure feed

heater train.



PAGE 63

FIG. 5.9 CONDENSER SUPPORT SPRING ARRANGEMENT

5.1.8.2 Circualting (Cooling) Water Leakage

There are two kinds of CW leakage, internal and external. Internal leakage into

the steam and condensate space is the most important of these two.

5.1.8.3 Internal Leakage

Leakage of cooling water into the condensate can be caused by several faults, but

the main ones are:

a) Tube to tube plate fixings leakage.

b) lnternalcorrosionanderosionofthetubes.

c) External erosion of the tubes.

d) Fatigue and stress cracking of the tubes.

The Effects of CW Leakage into the Condensate

Leakage of cooling water into the condenser steam side can have serious

consequences. The CW carries impurities with it, into the condensate system; the

most deterimental are those containing chlorides, such as sodium chloride (NaCI).

These impurities are then carried forward into the boiler.

The presence of chlorides in the boiler water constitutes a potential hazard,

principally because acid chlorides can be formed and boiler tube erosion can

result. The higher the boiler/pressure the greater is the danger. It is, therefore.

very important that CW leakage should be detected., the source of leakage located,

and the leak rectified.



PAGE 64

The Initial Indication of a Tube Leak

Fortunately, the impure water has a property which can be utilized to detect it. The

impure water conducts electricity betterthan the pure condensate and is said to have a

higher conductuvity. If the conductivity of the condensate is monitored a change will

be detected when a leakage of CW occurs.

The practical advantage of condensate conductivity measurement is that ft

indicates changes, not only in the actual value but also it increases above the normal

running value. To a plant operator, this often gives the first indication of condenser

leakage.

Location the Leaking Tube

There are several methods of locating the leaking tube, and new methods are

continually being tried. The principal methods of leak location are as follows:

a) The Simple Manometer. (Fig.5.10) shows a simple manometer, which can be

manufactured by the station chemist. One end of the condenser tube is plugged and

the manometer is inserted into the other end. The leaking tube will suck the liquid out

of glass because of the vacuum in the condenser. This method is ver ,y effective, but

can be time consuming.

b) The Blanket Effect. In this method the tube plate is covered by thin plastic

sheeting or by foam. The leaking tube will tend to pul! the foam or sheet into R.

c) Sonic Detection. As air is drawn into the leaking tube it creates a supersonic

whistle. This whistle is detected by a microphone placed in the entrance to the tube,

and the resulting signal is amplified. d) Bubbler Leak Detectors. (Fig.5.1 1) shows a

conventional bubbler. This can be used for on-load detection and its method of

operation is similar to the simple manometer.

If one end of the tube is blocked with a bung, the vacuum of the condenser pulls

air through the bubbler when it is inserted in the other end of the tube.

Amore advanced type of bubbler called MEL bubbler can be used both f or on-

load and off-load detection and is very suitable for use with pannier and integral

condensers.

(Fig.5.12) illustrates the principle of the MEL bubbler, The tube to be tested is

plugged, connected to a reference vessel and vacuum pump, and there the system is



evacuated. The pump is then isolated from the system and, after a short time, to

ensure that the pressure in the tube being tested and the reference vessel have

equalized, the balance valve is closed. From then on any air leaking into the tube

under test will be indicated by a stream of bubbles issuing from the end of the tube in

the jar or water.

PAGE 65

FIG. 5.14 (A) SACRIFICIAL ANODE TYPE OF FIG. 5.14 (B) IMPRESSED CIRCUIT SYSTEM OF



CATHODIC PROTECTION CATHODIC PROTECTION

If on-load location is not successful it may be necessary to take the turbine

off load to locate the leak. On some older turbines it is possible;to enter the steam

space and locate the leak directly as the CW sprays into the steam space, but

usually this is not possible on a modern turbine.

Whereundersiungcondensersarefittedftispossibletofilithesteamspacewithcon

densate.contain'ing fluorescene and examine the tube plate with an uftra-violet

lamp to find the leaking tube.

As we mentioned previously, the condenser support springs must be jacked

up be forelhe steam space is filled with condensate.

Where pannier or integral condensers have been fitted this method cannot

be used if the steam space is filled with water the low pressure cylinder could be

under water.

The Double-Tube Plate

On modern turbine, plant extensive use is being made of the double-tube plate in

an attempt ot reduce the effect of leakage at tube fixings. ( Fig. 5.13) illustrates

the priciple of the double-tube plate.

The interspace A can either be under vacuum (@,n which case leakage will

be into space) or ft can be fed with condensate under pressure a leakage from the

system. Alternatively, the conductivity of the drainage from the interspace A can

be monkored; an increase indicating a leaking tube fixing.

PAGE 66

5.1.8.4 External Leakage

External Leakage from condenser water boxes and joints is usually due to metal

removal by erosion or corrosion. Erosion is the physical removal of metal by

excessively turbui >entwater (particularly when it contains air bubbles), or by

water carrying grit or other suspended solids. This makes particularly susceptible

those places where water has to change direction quickly, such as water boxes, or

in areas of excessive turbulence due to the throttling action of valves. Leakage

path erosion between the impetier eye and casing of large C.W pumps may

necessitate the use of wearing rings at this point. An external leakage source may

also @ a broken anode in a cathodic protection system.



Corrosion is the result of electro-chemical action, which can be reduced but

cannot be entirely eliminated. Cast iron condenser waterboxes are particularly

affected by sea water, which dissolves the iron content of metal, leaving behind

weak and porous graphite in original shape. The application of protective coatings

and cathodic protection adoption help to reduce electrolytic corrosion. Painting

gives some protection to condenser water-boxes, although adequate surface

preparation and coverage is difficult to achieve, severe localised corrosion may

occur where there is a defect in point film. Natural or synthetic rubber coatings

are more successful and have a longer life, although initial cost is high. An

unprotected water box, however, provides some protection for copper alloy

condenser tubes by limited cathodic protection mechanism. Conversely,

successful coating of waterboxes accelerates corrosion, elsewhere, particularly at

tube ends. Thus, cooling should extend a short distance into tubes, or plastic

inserts may be placed in tube ends.

Cathodic protection is based on the principle of a corrosion cell; if two

dissimilar metals are placed in electrolyte, corrosion of the more electro-negative

one (anode) takes place in preference to the other (cathode). In cooling water

systems the iron components from the anodes and the copper alloys Oubes) from

the cathodes. If a third electrode, more electro-negative than the iron and the

copper alloys, is added to the system and is electrically connected to the other two

electrodes, the new electrode corrodes in preference to the iron, or the copper

alloys. This system is known as the sacrificial anode type of cathodic protection

(Fig.5.14) (a)), as the third electrode is sacrificed to save the original (iron)

electrodeimproved cathodic protection, requiring less maintenance (replacement of

sacrificial anodes) and able to protect a larger area, is provided by an impressed

circuit type system. It uses the sarm principle, but an inert or semi-inert material

(e.g. platinum coated titanium) is deliberately made anodic to existing material by

passing low voltage direct current through it into the electrolyte (Fig.5.14) (b)

5.2 CONDENSATE EXTRACTION PUMPS

Condensate extraction pumps are normally multistage, vertical,,

contrifugal,pumps (Fig.5.15). They are to generally required to operate on



minimum net positive suction head (NPSH). The condensate pumps operate on

few inches of suction submergence.

A vent line connects the hotwelf, from where the condensate pumps

take,suction with the condenser. This equalises the vapour pressure of condenser

and hotwell.

No. of stages in the pump is determined by the discharge pressure re uired

for the condensaite cycle.

In 21 0 MW unit, three condensate pumps, each having SO%- capacity, are

provided for pumping the condensate to deaerator. Condensate water is also used

for:

i) Sealing of glands of valves operating under vacuum.

ii) Temperature control of L.P. bypass steam.

iii) Filling syphons of main,ejectors and 15 meter syphon of drain expander.

iv) Actuating the forced closing non-return valves of turbine steam extraction

lines.

v) Operation of group protection device for bypassing H.P. heaters.

vi) For cooling steam dumped through.steam throw off device.

PAGE 67



LEGEND

1. DISCHARGE HEAD

2. OUTER SHELL

3. IMPELLER SHFT

4. IMPELLER FIRST STAGE

5. BELL MOUTH

6. STUFFING BOX HOUSING

7. THRUST BEARING HOUSING

8. UPPER PUMP BEARING

9. LOWER PUMP BRARING

10. GLAND IN TWO HALVES

11. ST. BOX PACKING

PAGE 68

Major specification of a typical BHRC 28 type condensate extraction pump (for 210 MW)

PUMP :

3 Nos. per unit Multistage, vertical turbine centrifugal pump. Low specific speed, medium head Medium Capacity Discharge - 281 T / hr. Manometric - 201 Metres. Head NPSH - 3.5 MTRS r.p.m. - 1489 No. of stages - 8 hp - 256

MOTOR

Power - 220 kW (300 hp), Voltage - 6.6 kV

5.3 AIR EXTRACTION SYSTEM

Air extraction system is needed to extract air and other non condensable gases from

the condenser for ma~kn vacuum.



Amount of air to be extracted from condenser during start up is quite large

and the extraction should be done as rapidly as possible so as to allow the turbine to

be started.

Under rkwmal operating conditions quantity of air to be extracted is lower. It

consists of air ieakage into the condenser via flanges and glands and also of very

little non condensable gases present in steam.

To guard against excessive water vapour extraction alongwith air, the space

beneath the air extraction baffles has been provided with its own cooling tubes in

order to condense as much water vapour as possible and thus preventing its removal

from condenser.

5.3.1 Air Ejectors

The oprerating medium of the air ejector can be either high pressur gas or liquid.

In thermal power stations steam of low parameter (Approx. 4.5 kg 1 cm', 250'c) is

used for the air e@or. The operating principle is simple - steam is passed through

a nozzle and the pressure energy converted into velocity energy. High velocity

fkjid aspirates air and other non condensable gases from the condenser and moves

into diffuser which re-converts the velocity energy into pressure energy@r The

pressu~ mixture of steam and air is exhausted, either directly to atmosphere or

through coolers to recover the steam in the form of condensate.

Starting Ejectors

Starting e@or is recommended to be used for accelerating the initial pulling of

vacuum. During this period starting e@or operates in parallel with main ejector.

When the vacuum in the condenser reaches 500-600 mm of Hq column, the

starting ejector is switched off.

(Fig. 5. 1 6) shows the construction of a starting ejector. It may be noted

that the steam alongwith the mbdure of air and other gases is exhausted to the

atmosphere. Gener ally starting ejector is single stage and has high steam

consumption.



PAGE 69

FIG. 5.16 AIR EJECTOR

Main Elector

The main ejector with a-standby unit is usually provided for normal operation.

The main ejector is a multi stage type, the number of stages depends on the

cooling water condition. Steam at suitable pressure is passed through a

converging-diverging nozzle and the pressure energy of steam is converted into

velocity energy. This high velocity steam jet entre fins air and in condensable



gases and then enters a diffuser steam lair mixture is ihen cooled in the first stage

shell by condensate. Steam is thus condensed, heat in the operating system is

partly recovered, and the steam 1 air mixture volume is reduced, allowing the

second stage nozzle and its steam consumption to be reduced. The second stage

cooler can be followed by a third stage nozzle, and its after cooler (as done in

BHEL 21 0 MW unit). Drains are usually returned to the condenser via suitable

loop seals; cooler condensate as a cooling medium is taken. from the extraction

pump discharge, with a recirculation arrangement to avoid overheating of the eject

at low loads. (Fig. 5.17) shows the arrangement of typical two-stage main ejector.

PAGE 70

FIG. 5.17 MAIN EJECTOR

An air measuring device for measurement of air discharge from condenser may

be fitted at the air exit of the ejector. It measures dry air discharge while the

condenser and ejectors are in operation.

5.3.2 Air Pumps In the ejector system high quality steam from the boiler is used by reducing both

pressur and temperature. For example, for running the ejection the main steam

from boiler having 140 kg cm2 and pressure 540 c temp. is reduced to 4.5 kg / cm

2and 250'c. Due to this reason the steam operated air ejector, which in other

circumstances is prefectly satisfactory, is not idealfor use inthe high pressure, high



temperature units. Hence, now-a-days air pumps are being used in the condensers

of the 500 MW units.

Air pumps allows greater flexibility as it is not dependent on the boiler for

raising vacuum. Air pumps operates on a separate water circuit and there is no

risk of the concentration of soluble un condensabie gases in the condensate. Air

pump can deal with either starting and normal conditions and there fore@a

separate starting equipment is unnecessary.

Air pumps are basically of 3 types : rotary, liquid ring and hydraulic.

a) Rotary Air Pumps (Fig. 5.18) It has two impeliers each consisting of a blanking plate mounted on the pump

shaft. The closely spaced blades are attached to the rim of the blanking plate.

Sealing, water is fed from an elevated tank in to the compartments

PAGE 71

formed by the blanking plate and the ends of pump casing.

When the pump rotates, water is drawn through the guide nozzle and is

broken into slugs by the impeller blades. These slugs of water pass the discharge

nozzle, aspirates the air and effectively seal the pockets of air asthey passthrough

the diffuser. Fromthe diffuser, the air, non-condensable gases and sealing water

tank. The swirling motion imparted to the mixture.by the tangential entry, assists

the separation of water and gases. The gases are released to atmosphere through a

vent in the top of the tank. The sealingwater is re-circulated to the pump.

b) Liquid Ring Type Air Pump (Nash Vacuum Pump) (Fig. 5.19).



There are several varieties of this type which is essentially a centrifugal

displacement type pump. A mufti blade rotor / impeller revolves within an

elliptical or offset casing which is partially filled with water. The rotating impeller

throws the liquid out wards resulting in a solid ring of liquid revolving in the

casing at the same speed as the rotor but following the shape of the casing. This a

alternatively causes the liquid to enter and recede from the inter blade spaces on

the impeller. The provision of inlet and outlet parts enables this pumping section

to be used for sucking in and out of the gases from the condenser.

Thesepumpsaresimpleandreliable.Therearelargeclearanceonrotatingpartsandnovalv

esorpistons,

c) Hydraulic Air Pumps : (Fig. 5.20)

Hydraulic air pump is a water operated ejector using a motor driven recirculating

lift pump and a simple spill over system. The sealing water is pressurised by a

piping pump and fed into a series of nozzles. Streams ol water leaving the nozzles

pass over the blades of spinner which is mounted on the end of a shaft and free to

rotate. The blades are formed so that the sealing water propels the spinner and is

broken into slugs. These slugs, like rotary air pump, pass to discharge nozzle,

aspirate air and seal the pockets of air in the diffuser and thus pressurises them and

discharges the mixture (of air and water) into a tank.

The tank is ven'ted and sealing water is recirculated to the lift pump.

PAGE 72



PAGE 73

B01LER FEED PUMP

Boiler Feed Pump (BFP) is a multistage pump provided for pumping feed water to

economiser. Generally three pumps each of 50% Of total capacity are provided,

For rated capacity two Pumps Will be working in parallel and the third will be in

reserve.

6.1 Description of Feed Pump (BHEL - 2oo KHI, provided for 200 / 21 0 MW Unit).

Feed pump consists of the following major parts:

1) Pump Barrel

2) Rotor

3) Stator

4) Mechanical Seal

5) Balanc7ing Device.

6.1.1 Pump Barrel

The barrel is essentially a cylinder which houses both the stator and rotor. The

suction side of the barrel and the space in the high press cover behind the

balancing device are closed by the low pressure covers alongwith the stuffing box

casings. The brackets of the radial bearing of the suction side and the bracket of

the radial and thrust bearings of the discharge side are fixed to the low pressure

covers. The entire pump is mounted on a foundation frame. As the pump handles

hot water, sometimes, arrangements are made for cooling the foundation frame to

prevent unequal expansion of the frame.

6.1.2 Rotor

The rotor of boiler feed pump consists of the shaft, impellers, distance bushes,

balancing disc, supporting rings etc. The axial thrust of the rotor is taken up by the

balancing disc. which is keyed to the shaft in between two the two parts

supporting rings which are mounted in the grooves in the shaft. The rotor is

supported on two part bearing shells. The bearing brackets are connected to the

low pressure cover.

6.1.3 Stator

The stator consists of stage bodies. The diffusers with the diffusing wheels and

guide wheels are assembled to the stage bodies. . The end diffuser is connected to



the outlet stage outside the stage body. Stage bodies are fitted with wearing rings

at the place where it is likely to come into contact with the wearing rings of

impellers, and the wearings rings are secured to the stage bodies woth the help of

screws.

6.1.4 Mechanical Seal

Sealing of the pump is achieved by a specially designed mechanical seal. It

minimises the loss of the feed water in the stuffing box and the working ability of

the pump increases. With the use of the mechanical seal, the cooling is carried out

by the circulation of water between the stuffing box space and the cooler. The feed

water is circulated in the cooling circuit through the cooler and back by means of a

pumping ring. The coolers are so designed that water temperature in the stuffing

box remains below 80o C..

6.1.5 Balancing Device

As in other multi stage pumps, all the six irnpellers are arranged on the shaft with

inlets in the same direction. This causes a thrust of about 34 Tons in the direction

of suction end of the pump while running. This axial thrust is taken up by the

balancing device. About 1 0% of feed water which is not calculated to the

guaranteed delivery

PAGE 74

PAGE 75



capacity is taken off from the space behind the last impeller for operation of the

automatic balancing device. The bakime disc is fixed to the shaft and rotates

between a ren~ie seating and the balance disc cover. The thrust generated by the

impeders tends to force the disc against its seating, but the high pressure water,

bled off the delivery stage of the pump, flows along an annular space between the

hub of the disc and the bush, which is an integral part of the balance disc seating,

to a pressure chamber.

The pressure in the chamber builds up until R exerts sufficients pressure on

the balance disc to overcome the end thrust of the impeller. Waterthen escapes

between the face of the disc and its seating. The balance disc thus runs on a film

of water and does not come into metallic contact with the seating. Water leakage

across this disc is called balance water and is returned to the deaerator.

A thrust kingsbe" bearing toes over the function of the balancing device when feed

pump is started. The & dngs berry shell is forced against the direction of action of

balancing disc on the disc by means of springs kwaed in the kings berry bearing. By

action of springs, an axial gap of about 1.0 mm is formed between the contract surface of

the bearing disc and balancing disc. The total pull of springs is equal to 500 kg. With the

starting of the pump the axial thrust increases gradually and the thrust kings berry bearing

is in action until the time when the magnitude of the axial thrust overcomes the pressure

of the springs mounted in mitchell bearings, the rotor Mil move to the suction side and

balancing disc comes into contact with bearing disc" reducing the a)dal gap and due to

the increased pressure on the balancing disc, the rotor move to the middle position

creating the gap between the balancing disc and the bearing ring.



Fig.62 DIAGRAM OF BALANCE PISTON ((DISC)

Even under worst condition when the rotor moves to the suction side and

the balancing disc is likely to come into contact with the bearing ring before the

necessary pressure being built up on the balancing disc to overcome the axial the

thrust, a certain amour* of water flows through the axial gap between the

balancing disc and the bearing ring and there is no danger of balancing device

getting seized.

PAGE 76

It is evident that behind the balancing disc the pressure must not rise.

Otherwise the hydraulic equalizing piping must have a sufficent flow capacity. For

safe operating, the pressure in the equalising piping should be 0.5 to 2 atm, higher

than the intake suction branch pressure. When the pressure in the balancing space

rises by 5 atm above suction pressure it is necessury to trip the pump in order to

find out the cause of defect and to rectify it.

6.2 WORKING OF BIOLER FEED PUMP

the water with the given operating temperature should flow to the pump under a

certain minimum pressure (NPSH)., water passes through the suction branch into the

intake spiral and from here is directed to the first impeller. After leaving through the

impeller it passes through the distributing passages of the diffuser where it get certain

pressure rise and flows over to guide vances to the inlet of the next impeller.

This process repeats from one stage to the other till it passes through the last

impeller and the end diffusers. Thus the feed water arriving into the discharge space

develops the necessary operating pressure, a small part of feed water i.e. about 10% is

taken off from the space behind the last impeller for the operating of the automatic

balancing device to balance the hydraulic axial thryst of the pump rotor.



TYPICAL SPECIFICATIONS OF BOILER FEED PUMP (200 KHI TYPE)

No. of stages 6

Suction Pressure 12.3 ata

Quantity of water for 100 Tons / hr.

Minimum take off

Discharge capacity / head 430 T / hr. / 1830MWC

Quantity of water for 8 Tons / hr

Warming up

Feed water temperature 164.2oC

Consumption of cooling 280 LPM

Water

Speed 4320 rpm

Lubrication Forces

Stuffing box Mech. Seal

Net weight of pump 5850 Kg.

Axial Thrust at Designed 34 Tonnes

Speed.

MOTOR

Output 4000 kW

Rated voltage 6.6 kV

Current 421 Amps.

Speed 1483 rpm

Frequency / Powe factor 50 c / s / 0.914

6.3 RECIRCULATION SYSTEM

To maintain a reasonable efficiency in the pump, running clearances between

stationary and rotating parts must be fairly narrow. Liquid flow through these

clearances acts as a lubricant to prevent seizure. The power input to the pump is

partly conveted into hydraulic energy due to the increase in pressure of the liquid.

The

PAGE 77



remaining energy is wasted in the form of friction, eddies and mechanical losses.

This power loss causes slight increase in the liquid temperature while the liquid

passes from suction to discharge. This temperature rise is maximum at zero

discharge and the water soon flashes into steam. Flashing breaks down the thin

film of lubricating water between the parts and this usually causes seizure. The

trouble occurs so quickly that stationary parts cannot expand as rapidly as the

rotating parts, because they will be heated more slowly, being of greater mass and

also exposed to atmosphere. Greater expansion of rotating parts will reduce the

normal running clearance and aggravate the conditions.

It is, therefore, imperative that sufficient water must be kept moving through the

pump to prevent its temperature from reaching the flash point in the pump when

the regulator closes the main discharge line due to low load or less water

requirements in the drum or when the pump is just started. To ensure this an

automatic leak off system is provided between the pump discharge and the

deaerator to establish a minimum flow through the pump. A solenoid operated

diaphragm valve or a motorised valve is installed in the leak off line which opens

when the pump runs at a lower capacity.

The recirculation valve (of BHEL 21 0 MW unit BFpp) opens when the flow at

pump suction is below 1 00 T 1 hr & closes when k increases to 220 T 1 hr. The

flow through recirculation line is 125 T / hr.

6.4 WARMING UP

Centrifugal pumps handling hot water should always be maintained nearly at

operating temperature when idle, If suddenly hotwater is admitted intothe pump,

the relative expansion of the casing barrel and of the inner element goes through

two separate phase. The inner elements expands faster than the barrel resulting in

distortion of the pump. To avoid this, asmall quantity of the medium is always

passed through the steam pump for warm up. Various methods are used for this

purpose. In some, the flow is from the suction, through the pump and out through

the balancing chamberto the flash tank. 1 n others, a by pass across the main

discharge non-retum valve is provided with a pressure reducing orifice. The flow

is from the discharge and, through the pump and back to the deaertor. Larger the



pump the longer is the time and the larger is the quantity of hot water required for

warming up.

In orderto avoid pump seizure in case of 200 KHI model while starting,

warming up valve should be open and temperature of pumps casing should not be

less by more than 150c than that of feed tank temperature.

6.5 VARIABLE SPEED HYDRAULIC (FLUID) COUPLING

Some boiler feed pumps including the KHI type are coupled with their driving

motor through a variable speed hydraulic coupling The hydraulic coupling serves

the purpose of controlling the speed of feed pump for maintaining definite delivery

head add delivered quantity of the feed water as per requirement of the boiler.

This reduces the power consumption particularly at part load operation.

6.5.1 Basic Principle and Operation

A fluid coupling is basically a combination of pump and turbine connected in

series. (Fig.6.3) shows avertable speed fluid coupling blading and the arrangement

of the impellers in the working area.

The rotating impeffer energy to the operating fluid. The resultant centrif

ugal force causes the fluid to flow outwards whereby the velocity is increased by

the impeller. The flow of the fluid into the runner takes place at the outer

diameter, where the energy is transmitted from the fluid. The fluid contained in

the runner blade chambers then flows inwards to the centre and back into the

impeller blade chambers. This circuit is maintained by the centrifugal force

difference resulting from the speed difference between the impetter and runner.

This speed difference is cabled slip, which nominally is in the order 1.5 to 3.5%.

In contrast to the constant-filled type turbo-coupling, the oil filling of the variable-

speed turbo coupling can be varied between fully filled and drained while in

operation. Ln this way stepiessspeed regulation of the driven machine over a large

range is achieved when the coupling operates against the load characteristics. This

regulating range is 4:1.

PAGE 78

The working circuit is governed by a system which can continuously extract or

supply the working compartment fluid. This enables precise adjustment of the

driven machine speed to be achieved.



The working circuit is charged by continuously running pump which delivers oil

from the int egral sump below the coupling into he working compartment. The

working compartment is the chamber between the primary (impeller) and

secondary (runner) wheels which is a connected to a rotating scoop chamber

consisting

FIG. : 6.3 VARIABLE SPEED FLUID COUPLING

PAGE 79

of an inner and an outer shell. The,oil level in the working compartment

deterrpines the speed at the output side of the coupling and depends upon the

radial position of ascoop tube located in the scoop chamber. The flow capacity of

the; scoop tube far exceeds the purnp delivery; thus, with respect to control and

regulation, reaction times are at a minimum.

6.6 BOOSTER PUMP (FIG. 6.4)

Booster pump is provided before feed pp. to maintain required NPSH and lower

Deaerator height. KHI type boiler feed pump is provided with a booster pump in

its suction line which is driven by the main motor of the boiler feed pump. One of

the major daimages which may occur to a B.F. pump is from cavaation or vapour

bounding at the pump suction 1due to suction failure. Cavitation will occur when

the suction pressure of the pump atthe pump suction is equal orvery



neartothevapour pressure of the liquidto be pumped at aparticular feed water

temperature. By the use oj-.booster pump in the main pump suction line, always

there will be positive suction pressure which will remove the possibility of

cavitation.

Typical Specification of Booster Pump Provide with 200 KHI Type BFPP.

Delivery quantity 460 T / hr.

Delivery head 31 MWC.

Speed 1475 rpm.

6.7 F.K. RANGE BOILER FEED PUMP

These pumps are barrel casing, multistage cart-ridge pumps. The cartridge

(Pump-Barrel) includes all - pump internals with shaft, impellers, Diffusers, shaft

seals, bearing housings and pump half coupling.. The cartridge design enables

easy removal of pump internals.

(Fig. 6.5) shows sectional arrangement of F.K. Type Boiler Feed Pump. lh

500 MW units in India, F3 4E K6 feed pump is used. Its specifications are given

below:

TYPICAL SPECIFICATIONS OF FK 4E 36 FEED PUMP

No. of stages 4 + 1 Kicker stage

Suction Flow 1080 M3 / HR

Head 2100 MTR

Speed 5690 RPM

Efficiency 81.75%

Power 6752 KW

Typical Arrangement 2 * 50 % Turbine Driven

1 * 50% Motor Driven

Interstage Tapping to supply water for Reheat Steam Desuperheating.

Kicker Stage Tapping to supply water for superheater Desuperheating.

Axial Thrust Balancing: Balancing Drum and Tilting Pad Thrust Bearing.

Axial Thrust at design speed: 72.5 Tonnes

Balancing Force at design speed: 66 Tonnes

Residual Axial thrust at de@igr4 speed : 6.5 Tonnes



Thrust Bearing Capacity : 14.5 Tonnes.

6.8 BFP DRIVE

For units of capacity upto 200 1 21 0 MW the drives for b ' oiler feed pumps are

electric motors. For units of capacity 500 MW and above the general practice is

to.employ steam turbines for driving the pumps. (An

PAGE 80

PAGE 82

electric motor driven feed Pump of 50% capacity is also provided for start up and

stand by). Number of steam turbines driven feed pumps per unit also vary. Some of

500 MW Units provide only one steam turbine driven feed pump of 1 00% capacity

and some provide two steam turbine driven feed pumps of 50% capacity each). It

normally takes steam from either CRH or lst Extraction and the exhaust normally

goes to one of the low pressure heaters.

One of the major gains in using steam turbine as the primemover for BFPs is

an increase in overall efficiency. The power consumed by feed pump is transfered to

feed water. But when the power consumed is electricity, ft has been generated at

around 33% efficiency only. So loss of power at feed pump drive is (in terms of %

of boiler heat input).

Pump Power (1 - n.)



Generator Load

Where nc = Cycle efficiency = nb x nt x ng

= 0.85 X 0.4 x 0.98 = 0.33 or 33%

nb = boiler efficiency = 0.85

nt = turbine efficiency = 0.4

ng = generator efficiency = 0.98

So if the power consumption of feed pump of a 500 MW Unit is 14 MW, then the

pump motor will consume 141 500 of the alternator output at an efficiency of 0.33

so, pump loss (% of boiler heat input)

= 14 - (100-0.33) / 500

= 1.88%

Now in case of turbine driven feed pump, the loss will be 14 / 500 (1 00 - n b) =

0.42% of boiler heat input. In addition, the steam turbine driven feed pump has

the foliowing other advantages:

1) Compared with the system using electrically driven pumps there is again in

therrnal and overall efficiency since thermally, the extraction of steam from the

feed pump turbine requires an increased steam flow through the early stages of

the turbine. This allows the use of longer blades resulting in higher stage

efficiency.

2) As the bled steam is normally taken from the cold reheat line the steam flow

through the reheater is reduced. This permits reduction of cost of reheater

steam pipe works because of the reduction of size. The main turbine

construction is simplified as there is no bled steam tapping points for HP

heaters.

3) Thermodynamically, it is advantageous to use the low temperature steam bled

from the feed pump steam turbine for feed heating in HP heaters as it helps

avoiding degradation of high super-heat steam tapped from H.P. and ].P.

turbine.



4) The feed pump speed is infinitely variable and it has been demonstrated that

the bled steam turbine will run the pump with the main turbine producing only

25% of full load.

5) There is no need for gearing between the turbine and pump.

6) The lower stage efficiency of the bled steam turbine is more than off set by the

higher transmission efficiency. The alternator shaft hydraulic coupling must

be designed such that the b.f.p. will give 1 00% output under frequency

reduction. Therefore under normal 50 Hz conditions the coupling will be

operating at less than its maximum range point with an increase in its losses.

7) A higher speed pump results in a smaller and cheaper pump. Also the b.f.p.

turbine operates at higher intenal efficiency at higher speeds. In order that high

pump-speed scan be obtained with an alternator shaft drive, a step-up gearbox

will be required, which increases the transmission losses.

8) Lt does not interfere with at ternator rotor with drawal and does not complicate

an are a already used for other auaxiliaries such as exciter and alternator

cooling fan.

9) Lt often allows a reduction in building costs as the over aillength of the T/A

unit is not increased.

PAGE 83



PAGE 84

10) The turbine driven feed pump may be sited in the basement instead of at

engine room floor level, so that for a given N.P.S.H. the deaerator height may

be lowerd.

(Fig. 6.6) shows a K 1401 type B.F.P. turbine used in 500 MW (KWU) Unit.

PAGE 85

REGENERATIVE FEED HEATING SYSTEM

7.1 Economics of feed heating

lf steam is bled from a turbine and is made to give up its latent and any supe.-heat it

may possess, to a heater, this system is called regenerative, because the fluid (steam)

gives up heat, which would be otherwise wasted, to the fluid whilst in another state

(water) to raise its temperature. The highest A heorectical temperature to which the

feed water may be raised in the heater is the saturation temperature of the bled steam.

There is an optimum point at which the steam is bled from the turbine once a feed

temperature is selected, a tappling point near the stop valve produces no gain in

efficiency as practically live steam is used for heating. An intermediate point, if

carefully chosen, gives maximum feed temperature rise with minimum loss of

mechanical power at the turbine. The steam, having given up a proportion of lits work

to the turbine, then gives up all its latent heat which would otherwise be lost to the

condenser C.W.The heat gained in this way outweighs the loss of mechanical power and

a gain in efficiency follows. Other advantages of this cycle are that less C.W. is

required with a decrease in pumping power, a smaller condenser can be used and the

turbine exhaust annulus is smaller.

The thermal gain resulting from feed water heating can be illustrated by

considering an example with approximate figures as follows (Single Feed

Heater).

Without Feed Heating With Feed Heating

Turbine steam consumption 4.5 kg / kWh 4.5 x 1.05 = 4.725 kg 1 kWh

Boiler feed temperature 31oC 93oc

Total heat in steam 314 x 4.5= 3140 x.725=



14130 kJ kwh 14836.5 kJ / kwh

Heat already in water 4.5 x 4.19 x 31 4.725 x 4.19 x 93

= 584.5 kJ / kwh = 1840 kJ / kwh

Heat used by turbine (4.5 x 3140) - 584.5 (4.725 x 3140) - 1840

Heat in mass of steam = 14130 - 584.5 = 14836-1841

- Heat to raise water (1 3545 kJ / Kwh) = 12995 kJ / kwh

The difference between these values is 4.23% which can result in a considerable

yearly saving in fuel consumption.

7.2 Types of feed water heaters

A feed water is simply a heat exchanger which is arranged so that the water leaving

a condenser is pre-heated before it is fed to a boiler. The feed heater is supplied by

steam which has already performed some useful work. This steam, which is taken

from suitable stages along a turbine, transfers its latent heat to the boiler feed water

and accordingly increases the water temperature.

It is now universal practice to use feed heaters to heat the feed water from the

temperature at which it leaves a condenser to a temperature approaching the

saturation temperature of the boiler steam pressure.

When a feed heater is in operation, it requires no regulation because the bled

steam consumption responds automatically to the temperature and quantity of feed

water passing through the heater.

Low pressure feed heaters are positioned after an extraction pump, while high

pressure feed heaters are positioned after a boiler feed pump and, therefore, have to

be constructed to withstand the full discharge pressure of a boiler feed pump.

Two types of feed heaters are used; the surface type, in which the feed water is

passed through tubes, with the bled steam surrounding them; and the direct contact

type, in which the steam and water mix together.

PAGE 86 7.2.1 The surface – Feed Heater

A surface type heater consists basically of three parts;

a) A shell with a steam connection

b) A nest of tubes



c) A water box, with water connections on the inlet and outlet headers.

A low pressure surface heater (Fig. 7.1) may consist of a cylindrical body

fabricated from mild steel and sealed at its upper end by a cast steel water box. Which

houses a nest of solid brass U-tubes. The ends of the tubes are expanded into a mild steel

tube plate trapped between a flange on the body and a corresponding flange on the water

box. Baffles are provided to ensure that the steam is directed across the tubes. The upper

section of the quadrant of the tube nest, which carries the condensate in its last pass

through the heater, is totally enclosed by vertical baffles, so forming a flashed-steam

drain cooler section of the heater.

FIG. 7.1 L.P. HEATER

PAGE 87

high pressure surface heater (FIg. 7.2) consists of a mild steel cylindrical

body and some secured together to form a vessel, which houses a tube nest of

carbon steel tubes suspended from a condensate inlet and outlet heater.



A typical modern high-pressure heater has integral desuper heating and

drain cooling zones in addition to condensing surfaces in the one shell. (Fig. 7.3)

shows a temperature / heat transferred diagram for this heater. The outlet

terminal difference is –0.56oC(-1oF) and the drain cooling terminal difference is

5.56oC

FIG. 7.2 H.P. HEATER

PAGE 88

7.2.2 Heat Transfer in Feed Water Heaters

There are three (3) heel transfer zones in a typical H.P. surface feed heater.

(a) esuperheating Zone

This zone is the @ feed header zone through which the feed-water passes

before leaving the heater. When the feed-weder enters the zone it has been heated

in the condensing, zone to within a few degrees of the saturation temperature,

corresponding to the bled steam pressure at the entry to the feed heater. Although

the desuperheating zone adds only a small percentage of the total heat transfered



in the heater, this small rise in feed-water temperature is very valuable in terms of

thermal economy. The heat in the zone is transferred by convection - the

superheater steam can be considered to behave as a normal gas.

The steam normally leaves the desuperheating zone with a residual

superheat of about 27.8'c. The temperature gradient between the bulk steam and

feed water temperature is shown in the top right hand inset of (Fig. 7.3). The tube

wall temperature is above the saturation temperature of the bled steam and no

condesnsation takes places on the tube wall under normal conditions. There is,

therefore, no problem of condensed steam forming droplets and being carried into

the condemning zone by the relatively high velocity steam which could cause

impingement attack @ the exit from the zone.

(b) Condensing Zone

In the condensing zone the overall coefficient of heat transfer is high. The

main problem in larger heaters is to obtain good steam distribution tothe

condensing surface with the minimum pressure loss. Under normal load running

conditions the ingress of non-condensabie gases is unlikely. However, continuous

venting is necessary to help the steam distribution and to clear any non

condensable gases which are likely to accumulate after shut down during two shift

operation. The diaphragm plates serve a dual purpose, not only do they support

the tube nest, but they also prevent the accumulation of condensate from running

down the outsiders of the tubes and forming a thick film with a consequent

reduction in heat transfer.

(c) Drain Cooling Zone

It is possible to obtain a fairly high beat transfer rate in the drain cooling

zone provided the drains are not reheated by head transfered through the shrouds

around the drain cooler convection zone. If there were. no desuperheating by a

steamy atmosphere which would tend to condense on the shrouds and reheat the

drain In the heater shown in Fig.7.3 the desuperheating zone is larger than the drain

cooling zone and it is possible to have the heater shell flooded with condensate to

above the drain cooler inlet . This condensate forms a stagnant pool around the

drain cooler shrouds causing them to follow the temperature gradient in the drain



cooler. Under this condition the problem of drain reheat is eliminated provided the

shell drain valves are kept closed.

7.3 REGENERATIVE SYSTEM OF 210 MW (BHEL LMW) UNIT

The regenerative system of the turbine consists of four low pressure (LP) heaters,

two gland coo ers, one deaerator and three high pressure (HP) heaters. The

condensate is drawn by condensate pumps from the hot well of condenser and is

pumped to the deaerator through gland cooler and low pressure heaters where it is

progressively headed up by steam extracted from sems and bled points of the

turbine. The drain of condensed steam on LP heaters No.2,3 and 4 flows in cascade

and is ultimately pumped into the main condensate line after heater No.2 orflowsto

condenser. The feed water after being deaerated in the deaerator is drawn by the

boiler feed pump and pun~ to boiler through high pressure heaters where it is heated

up by the bled steam from the turbine. The drain of condensed steam of HP heaters

flows in cascade and under normal load conditions flows to the deaerator.

PAGE 89

FIG. 7.3 TEMPERATURE / HEAT TRANSFERRED HIGH-PRESSURE HEATER

7.3.1 Low Pressure Heater No. 1

The heater is of horizontal surface type consisting of two halves, each half has

been located inside the upper part of each condenser. The two halves have been

installed.in parallel. The steam to both is supplied from the same extraction point.



The housing forthe heaterisfabricated from M.S. Plates with suitable steam

inlet and drain connections. The tube plate is of mild steel and is secured to the

water box and housing by means of studs and nuts. U. shaped tubes have been

used to ensure independent expansion of tubes and the shell. They are of solid

drawn admirality brass, 19 mm external dia, 1 mm and 0.75 nwn thick and are

expanded by rolling into the tube plate at facilitate drawal for tube replacement,

and maintenance. Partitions mild steel plates have been provided for supporting

the tubes at intermediate points and effective distribution of heat load in all the

zones of the heater.

The water box is of mild steel with suitable water inlet and outlet branches.

It is of rectangular shapes and has been provided with suiitable air vent and drain

connections.

7.3.2 Low Pressure Heater Nos. 2,3 & 4 (Fig. 7.4).

a) Construction

These heaters-identical in construction are of vertical surface type and are

designed for the steam to pass over the tubes and the condensate to flow

through them. Following are main elements of these heaters.

i) shell

ii) tube system

iii) removable water box.



PAGE 90

FIG. 7.4 L.P. HEATER 2,3 & 4 (210 MW 1 LMW UNIT)

Shell is a cylindrical construction with dished end welded at bottom and

having a flange at the upper end for assembly of tube system and water box. The

shell is provided with suitable steam inlet and drain connections alongwith other

nozzle connections to accommodate various fittings. M.S. baffles are provided to

ensure effective distribution of steam in the condensing zone of the heater.

Tube system consists of U-shaped admirality brass tube, 16 mm external

dia, 1 mm thick and are expanded by rolling into tube plate at both the ends.

Tube system has been provided with rollers to facilitate drawal for tube

replacement. Tube plate is of mild steel and is secured to the water box and shell

flange by means of studs and nuts.

Water box consists of thick walled cylindrical shell having a flange at the

lower end and a dished end welded at top. It has been provided with suitable

water inlet and out let branches. Partitions have been provided in the water box to

make it four path design.

PAGE 91

b) Working Principle



The main condensate f . lows through the tubes in four Paths before leaving

t he heater. The heating steam enters the shell through a pipe and flows over the U

shaped tube nest. The partition walls installed in the tube system ensures zig-zag

flow of steam over tube nest. Condensate of heating steam referred as drain,

trickles down the tubes and it taken out from the lower portion of the shell by

automatic level control valve installed on the drain line.

7.3.3 Gland Steam Cooler No. 1

It cools the air-steam mixture sucked from turbine end seals. It is of vertical type

and has two sections. An ejector mounted on the cooler, maintains constant

vacuum in the first section. It also sucks the remaining air steam mixture from ist

section to second, where air is let off and steam condensed., A part of main

condensate, after main ejector, flows through the cooler tubes consisting of U-

shaped brass tubes rolled in steel tube plate. Drain from cooler is led to

condenser.

7.3.4 Gland Cooler NO.2

Gland cooler has been designed to condensate the leak-off steam from

intermediate chambers of end sealings of HP & IP turbines.

The construction of this cooler is identical with low pressure heaters No.

2,3 & 4.

The main condensate flows through the tubes in four paths before leaving

the cooler. The leak off steam enters the shell through a pipe and flow over the

tube nest . The participation walls installed in the tube system lead to zig-zag flow

of steam over the tube nest. Condensate ofleak off steam referred as drain trickles

down the tubes and is taken out from the lower portion of the shell by automatic

level control valve, installed on the drain fine.

7.4 DEAERATOR

7.4.1 Functions

The pressure of certain gases like Oxygen, carbon dioxide and amonia,

dissolved in water is harmful because of their corrosive attack on metals,

particularly at elevated, temperatures. Thus in modern high pressure boiler, to

prevent in ternal corrosion, the feed water should be free, as far as practicable, of

aildissolvedgases, especially oxygen. This is achieved by embodying into the



freed system a deaeratifig unit. Apart from this, a deaerator also serves the

following functions:

1) Heating incoming feed water.

2) To act as a reservior to provide a sudden or instantaneous demand.

7.4.2 Principle of Deaeration

a) The solublity of any gas dissolved in a liquid is direct typroportional to the

part ial pressure of the gas.This holds within close limits for any gas which

does not react chemically with the solvent.

b) Solubility of gases decrease with increase in solution temperature and or

decrease in pressure.

7.4.3 210 MW LMW Unit Deaerator (Fig. 7.5a)

A constant pressure deaerator, pegged at 7 kg/ cM2 (abs) is provided in turbine

regenerative cycle to provide properly deaerated feed water for boiler, limiting

gases (mainly oxygen) to 0.005 cc/ Litre. It is a direct contact

PAGE 92

PAGE 93

type heater combined with feed storage tank of adequate capacity. The heating

steam is normally supplied from turbine extractions but during starting and low

load operation the steam is supplied from auxiliary source.

The deaerator comprises of two chambers:

i) Deaerating column



ii) Feed storage tank

Deaerating column is a spray cum tray type cylinderical vessel of horizontal

c ' onstruction with dished ends welded to it. The tray stack is designed to ensure

maximum contact time as well as optimum scrubbing of condensateto achieve

efficient deaeration. The deaerating column is mounted on thefeed

storagetankwhich in turn is supported on rollers at the two ends and a fixed

support at the centre. The feed storage tank is fabricated from boiler quality steel

plates. Manholes are provided on dearating column as well as on feed storage

tank for inspection and maintenance.

The feed water is admitted at the top of the deaerating column and flows

downwards through the spray valves and trays, The trays are designed to expose to

the maximum water surface for efficient scrubbing to effect the liberation of the

associated gases. Steam enters from the underneath of the tray and flows in

counter direction of condensate. While flowing upwards through the trays,

scrubbing and heating is done. Thus the liberated gases move upwards alongwith

the steam. Steam gets condensed above the trays and in turn heat the condensate.

Liberated gases escape to atmosphere from the orifice opening meant for ft. This

opening is provided with a number of deflectors to minimise the loss of steam.

In some deaerator designs, a vent condenser (Fig.7.5b) is also located above

the Deaerator. A portion of feed water is first passed through the vent condenser

before it enters the Deaerator. This water is heated by remaining steam after

steam has passed through the Deaerator. Thus only gases escape to atmosphere.



7.4.4 Location of Deaerator

A deaerator is placed at a height of about 20 Mts above B. F.P suction to avoid

flashing and cavitation during a rapid load drop.

PAGE 94

During a rapid load drop pressure in the heaters and deaerator tends to drop. This

causes flashing in the deaerator as the water is stored at boiling point,

corresponding to the pressure at full-load. The rate at ic head as it descends must

be greater than the rate of the water in the feed pump suction pipe gains stat'

pressure decay in the deaerator, if flashing in the pump is to be avoided. The

suction pipe should be as near vertical as possible to avoid unnecessary head loss

duetofriction.(Fig.7.6) illustratesthe variation in pressure in a deaerator and at a

feed pump positioned beneath it follomng a load rejection. The curve of pump

suction pressure is greater, by a constant figure than the curve of deaerator

pressure; this constant difference is equal to the static head of the deaerator on the

pump. If the saturation pressure corresponding to the ternperadure of the water

arriving ad the pump is greaterthan the total pressure atthe pump inlet, in other

words static head plus deaerator pressure at that time less friction loss, then



flashing will take place in the pump. because of static head, this situation is

avoided. In Modern units, a booster pp-is located before the Main feed pump to?

further increase the feed pp-suction pressure above saturation pressure at feed

pump suction.

FIG. 7.6 PRESSURE CHANGES AT A DEAERATOR AND FEED PUMP

FOLLOWING LOAD REJECTION

7.5 HIGH PRESSURE HEATERS (Fig. 7.7)

The feed water flows through the tube spirals and is heated by bied steam around

the tubes in the shell of the heaters. These heaters are cylindrical vessels with

welded dished ends and with integrated, desuperheating, condensing and

subcooling sections. The internal tube system of spirals is welded to the inlet and

outlet headers. As there are no flange ends the chances of tube leakages are less in

this type of heaters. In order to facillitate assembly and disassembly, rollers at the

side of the heater have been provided. Both feed water and steam entries and exits

are from the bottom end of the heater.

In 21 0 MWILMW units, the feed water, after feed pump enters the HPHs

5,6 & 7. The steam is supplied to these heaters from the bleed point Nos.3,2 & 1

of turbine through motor operated valves. These heaters have a group bypass

protectionon the feed water side, in the event of tube rupture in any of MPHs and

the level of the condensate rising to dangerous level, the group protection device

diverts automatically the feed water directly to boiler, thus by passing all the 3

H.P. heaters.



The condensate of the bled steam formed in the heater is thrown either to

the next lower stage heater in cascade or to the deaerator through a set of inter-

locked valves depending upon the pressure condition inside the heaters. There is

also an arrangement to take out;dir steam mixture from each heater in cascade and

air steam mbdure is thrown to the condenser through the LP heaters.

PAGE 95

PAGE 96

TURBINE OIL SYSTEM

8.1 PURPOSE OF OIL SYSTEM

The turbine oil system fulfils four fuctions. It:

a) Provides a supply of oil to the journal bearings to give an oil wedge at the shaft

rotates.

b) Maintains the temperature of the turbine bearings constant at the required level.

The oil does this by removing the heat which is produced by the shaft

conduction, the surface friction and the turbulence set up in the oil.

c) Provides a medium for hydraulically operating the governor gear and controlling

the steam admission valves.



d) Provides for hydrogen-cooled generators a sealing medium to prevent hydrogen

leaking out along the shaft.

It is worth noting that for 500 MW unites and above, h is becoming the

practice to use fire resistant fluids in place of lubricating oil for the control of

governer gear and steam admission valves. These eliminate the risk of fire

caused by leakage which is particularly likely when higher fluid pressures are

used.

(Fig.8. 1) shows the schematic of the lubricating oil system for 21 0 MW

(BHEL/ LMW) turbine. Lubricating oil systems for power station steam turbines

of other make or ratings also are a more or less simillar.

OIL SPECIFICATION (210 MWJLMW Turbine 011)

1 . Recommended Oil

a) Turbine oil 14

b) Mobil DTE medium

2. a) Specific gravity at 50'c 0.852

b) Kinematic viscosity at 50'c 28 CS

c) Neutralisation number 0.2

d) Flash point 201'c (min)

e) Pour point -6.60c (max)

f) Ash percentage by weight 0.01%

g) Mechanical impurities Nil

8.2 SYSTEM

The turbine oil system consists of the following:

1 . Main oil pump

2. Starting oil pump

3. A.C. Lub oil pump

4.. D.C. emergency oil pump

5. Oil tank

6. Drain valve

7. Oil pressure drop relay (OPDR)

8. Oil Coolers



Lubricating oil is supplied from oil tank (capacity 28,000 lftres) to bearings

and governing system with the help of pumps.

During period of normal operation the required oil is supplied through the

main oil pump mounted on the turbine shaft. A portion of discharge of the main

oil pump is used as the working oil for the injectors. In fact there are two injectors

located in the oil tank. The first injector supplies oil to the suction of the main oil

pump and the discharged oil is further pressurised through the second injector

which supplies oil to the bearings through coolers. PAGE 97

PAGE 98

During initial starting a A.C. driven starting oil pump meets the requirement of

both the bearing oil and governing oil.

Two standby oil pumps are incorporated in the system to supply bearing oil in

emergency. One of these is A.C. driven and the other is D.C. driven.



8.2.1 Main Oil Pump

This pump is mounted in the front bearing pedestal 'It is coupled with turbine rotor

through a gear coupling. When the turbine is running at normal speed i.e. 3000

rpm or the turbine speed is more than 2800 rpm, then the desired quantity of oil to

the governing system at 20 Kg 1 cm' (gauge) and to the lubrication system at 1 Kg

1 CM2 (gauge) is supplied by this oil pump. The oil to the lubrication system at

the level of turbine axis is supplied through two injectors arranged in series. First

injector develops a pressure of 3 Kg 1 CM2 (gauge) before oil coolers. After the

oil coolers, the oil pressure is 1 Kg 1 CM2 (gauge) which goes to lubrication

system. (Fig. 8.2) shows main oil pump of a 500 MW unit. It is similar to main oil

pump of 21 0 MW unit.

8.2.2 Starting Oil Pump (Auxiliary Oil Pump)

It is a multi-stage centrifugal oil pump driven by A.C. electric motor. Starting oil

pump is provided for meeting the requirement of oil of the turbo set during

starting. During starting or when the turbine is running at a speed lower than 2800

rpm ft supplies oil to governing system as well as to the lubrication system.

PAGE 99

8.2.3 A.C. Lub Oil Pump



This is a centrifugal pump, driven by an A.C. electric motor. This runs for about 1

0 minutes in the beginning to remove air from the governing system and to fill the

oil system with the oil. This pump automati over under inter lock action whenever

the oil pressure in lubrication system fails to 0.6 kg 1 CM2s (guage). Thus

8.2.4 D.C. Emergency Oil Pump

This is a centrifugal pump, driven by D.C. electric motor. This pump has been

provided as a back-up protection to A. C. driven lub. oil pump. This automatically

cuts in whenever there is failure of A.C. supply at power station and or the

pressure in the lubrication system fails to 0.5 kg 1 cm, (gauge).

8.2.5 Oil Tank

The oil is stored in oil tank of 28000 litres capacity upto operating level of the

tank. About 4000 lit / min. oil remains in circulation. Liberally sized tank holds

the oil inside the tank for a period long enough to ensure liberation of air from the

oil. Different mesh sized fitters are located inside the tank to filter the oil during

its no~ course. The filters are easily accessible and removable for cleaning even

when turbine is in service. This oil tank is supported on the framed structure just

below the turbine floor at the left side of the turbine.

8.2.6 Relief 1 Drain Valve

This valve is mounted on the oil pipe fine to maintain 1 kg / CM2 (gauge)

pressure at bearing axis in the lubrication system. So, whenever the pressure in

lubrication system differs from the above value, the drain valve increases or

decreases the oil drain to the tank and thereby maintain the required oil pressure in

the bearing lubrication system.

8.2.7 Oil Pressure Drop Relay (OPDR)

The A.C. lub. oil pump and emergency oil pump come in service automatically

under interlock action. The impulse for bringing the pumps into service is

provided by oil pressure drop relay. It is possible to over ride the interlock

manually to bring both pumps in service.

This OPDR comes into action under following conditions:

a) lt switches on the A.C. lub. Oil pump when the pressure in the lubricating

line drop stoO.6k91CM2(gauge).



b) It switches on the emergency oil pump when the pressure in the lubricating

jine drops to 0.5 kg / cm (gauge).

c) It trips the turbine and prevents the operation of barringge arwhen lubricating

oil pressured ropstoo.3 kg 1 cm' (gauge).

The relay is provided with a drain valve and an isolating valve for testing the

reliability of as operation, even when the turbine is running.

8.3 TURBINE OIL PURIFIER (CENTRIFUGE) The purifier is used for purification of turbine oil. It draws oil from the turbine oil tank

(or impure oil. It draws oil from the turbine oil tank (or impure oil tank located outside

turbine building) through a oil pump. After removing any water and entrained solid

matter, the clean oil is returned to the oil tank (or pure oil tank).PAGE 100

The Centrifuge can perform the following Functions 1. Clarification : A liquid (oil) – Slude separation in which the machine is used for

separating off particles normally solids.

2. Purification : Liquid (oil) – Liquid (water) separation in which the machine is

used for separating two intermixed liquids which are insoluble in each other and a

have different specific gravities solids with specific gravites higher than those of

the liquids that can be separated off at the same time.

FIG. 8.3 CENTRIFUGE

PAGE 101

8.3.1 Centrifuge

The centrifuge (Fig.8.3) consists of a bowl which rotates on a vertical axis in

an,outer casing. A 415 V induction] motor, mounted on the side of the casing

drives the bowl through a. centrifugal clutch and a worm gear; a manually



operated brake is included on the worm shaft. The outer casing has a hinged lid

which incorporates the oil inlet and outlet connections, indications forflow and

temperature and a priming water supply, The base, which has an oil sump in which

the worm gear runs, is mounted on the purifier.

The bowl, rotated by a vertical shaft, consists of a body and hood joined

with a locking ring. The body contains a stack of conical discs on a distributor,

the top disc has a level ring which bridges the hub of the distributor, immediately

above the level ring is fitted a paring disc. The dirty oil inlet passes through the

bowl hood gravity ring and paring disc to the hollow distributor.

8.3.2 Principle of Operation

The mixture of,dirty oil, water and solid impurities, flowsthrough ducts in the inlet

connection on the centrifuge hood and through ports in the walls of the inlet pipe

into the spa 1 ce between the hub of the bowl and the bore of the distributor. The

mixture flows through holes in the conical base of the distributor and the discs to

the inter spaces between the revolving discs where the water and solids are thrown

outwards by the centrifugal force and so are separated from the oil. The water and

solids pass along the undersurface of the discs towards the edge of the bowl where

the solids are deposited and so can be removed later when the bowl is dismantled

for maintenance.The water rises above the top disc and is discharged through the

annular space between the gravity disc and the inlet connection into the outer

casing and then out through the inspection box.

The clean oil goes along the upper surface of the discs towards the centre of

the bowl and is discharged through the bore of the level ring tothe paring disc.

The oil is kept in the centre of the bowl by the water forming a rotating seal: This

prevents the oil from flowing over the rim of the top disc and escaping through the

gravity disc. There is, however, a small region where the oil and water meet

known as the inter-face. The position of this interface can be varied by altering

the size of the bore of the gravity disc and so altering the flow of the sealing water

out of the bowl. The position of the interface also determines the purity of the

water and the oil, A smaller bore in the disc will move the interface towards the

centre of the bowl so that the water will contain less oil and vice versa. It should

be noted that if the specific gravity of the turbine lubricating oil changes it will



also affect the interface and the gravity disc will have to be changed. The water

seal has to be established on start-up by filling the bowl with hot water; this can be

supplied from the pump through the inlet connection on the hood.

The clean oil leaving the bowl rotates in the form of a ring around the

paring disc which is fitted with internal radial scoops. These scoops dip into the

liquid ring converting its kinetic energy into pressure to pump the oil through the

annular space, between the oil inlet tube and the housing, to the outlet connection.

8.3.3 Anti-flood Arrangements and Drains

The water seal is essential for the centrifuge to be operated correctly; if the seal is

broken (because of a low water contant in the oil and evaporation or faulty

operation), oil will be discharged through the.,water drain. The water sump in the

centrifuge mounting platform is divided into two sections by a low weir plate ane

section is occupied by the floats of two mercurys witches mounted on the

platform, When operated normally, the water separated from the oil is discharged

into the other section of the sump and drains through a discharge valve which is

set to pass the expected flow.

If oil is discharged with the water, the sump will overflow into the section

containing the floats; these rise and operate the switches. This automatically stops

the centrifuge motor.

PAGE 102

HP-LP BYPASS SYSTEM 9.1 INTRODUCTION

This bypass system has been provided to allow the steam generator to build up,

during start up, matching steam parameter with the turbine. The steam generated

is dumped into the condenser, thus avoiding loss of boiler water. This system

enables starting of the unit of sliding parameters and also facilitates hot restarting

of the unit. In the event of loss of load on the turbine, the bypass system disposes

the steam produced by the boiler automatically to the condenser without affecting

the boiler operation.

The bypass system has two section: HP & LP. The HP-Bypass system

diverts the steam before main steam valve (MSV) to the cold reheat CRH line. HP

Bypass system also reduces the rated steam parameters of the incoming steam



from the superheater (SH) to the steam conditions expected in the CRH line (i.e.

steam temp. and pressure after HP Turbine exhaust).

The LP Bypass system diverts the incoming steam f rom hot reheat (HRH)

line before intercepting valves (]V) to the condenser after reducing the HRH steam

parameters to the condition approximately to that of LP steam turbine exhaust

steam (Fig. 9.1).

HP Bypass station is utilised for the following tasks.

i) To establish flow at the outlet of superheater (SH) for raising boiler

parameters during start up.

ii) To maintain or control steam pressure at preset value in main steam line

during start up.

iii) To warm up the steam lines.

iv) To control steam temperature down stream of HP bypass at the preset value.

v) To dump steam from boiler into condenser, via LP-Bypass system, in case the

generator circuit breaker opens.

LP Bypass station is utilised for the following tasks.

i) Control of steam pressure after reheater.

ii) Establish flow of steam from reheat lines to condenser by its opening,

proportional to the opening of HP bypass valves.

iii) Release of steam entrapped in HPT and reheater circuit in case generator

circuit breaker opens.

Feed water is used for cooling H.P. bypass station and condensate water is

used for cooling of L.P. bypass station, condensate is also used for steam

cooling of H.P. 1 L.P. Bypass valves.

9.2 DESIGN CAPACITY OF H.P. BYPASS STATION 210 MW (LMW) UNIT

Total capacity of both bypass valves = 2x100= 200T/hr.

Max. M.S. Temp. = 5400c

Max. M.S. Pressure = 140 ata.

Down stream temperature = 380'c (Max.)

In KWU design, the capacity of HP & LP Bypass valves varies from 30% to 1

00% as per requirements of different customers. Higher capacity bypass valves

involve high cost.



9.3 CONTROL

a) The control components are located in a control cabinet in unit control board

(UCB). Each positioning loop may be controlled separately from the central

control desk. For supervision of the control loops, the position and control

deviation are indicated on mosaic insert of the control desk.

b) 0i1 Supply Unit : The oil supply units for the high pressure bypass and the

low pressure bypass are connected in parallel. Monostats control the oil

pressure. in the accumulators and signal alarm "PRESSURE TOO HIGH" or

"PRESSURE TOO LOW" appear in UCB if the pressure is not in order. If the

oil pressure should fall below the minimum in both accumulators, positioning

actuators will be blocked, and thereon the signal "ACTUATOR BLOCKED"

shall appear in UCB which simultaneously changes the operation of

positioning loops from automatic to manual.

PAGE 103

PAGE 104

9.4 INTERLOCKS

1 . The HP-bypass system are influenced by the following interlocks:

a) Generator circuit - breaker



b) Condenser vacuum too low

c) HP valve position

d) Temperature too high at down stream of LP bypass station.

Following interlocks are produced by the bypass system and given to the other

positioning loops.

e) HP valve position more than 2%

f) “Close" - signal for spray water pressure control valve.

2. Generator Circuit Breaker: The HP bypass station shall come in to operation at

the moment the logic signal "GENERATOR CIRCUITBREAKER CLOSED"

disappears. However, this does not implythatthe HP bypass station shall be

out of service if the logic signal "GENERATOR CIRCUIT BREAKER

CLOSED".

3. Condenser Vacuum Too Low : The HP bypass station shall close immediately

in case of too low condenser vacuum (500 mm Hg Cot.) This interlock holds a

first priority for the controller.

4. HP Bypass valve position less than 2%:

a) When turbine is running or not running and the control of HP bypass valve is

on manual, the memory will get closing signal through AND logic, 9 the valve

position is less than 2%.

b) When turbine is running and control of HP bypass valve is on auto, the memory

will get closing signal through AND logic, if valve position is less than 2%.

c) When turbine is not running and control of HP bypass valve are on auto, there

is no closing signal to memory whatever is the position of the valve and thus

pressure control loop will actuate the valve.

5. Temperature too high: if the temperature after the outlet of the HP or LP bypass

station becomes "TOO HIGH" the closing signal to HP bypass valve is

forwarded and simultaneously positioning loops changes from automatic to

manual mode. The inter lock for this case will be provided by the temperature

supervising monitor.

6. HP bypass valve position more than 2%: if any of the HP bypass valves are

opened more than 2% (which is initiated by a part of voltage monitor) or if the

position demand signal is equivalent to more than 2%, valve opening (which is



initiated by a part of voltage monitor) a signal is available through or logic to

indicate the valve is open and the same signal is used to change the control of

valves from manual to auto if their control was on manual.

9.5 DESIGN CAPACITY OF L.P BYPASS STATION : (210 MW, LMW UNIT)

Total capacity of both bypass valves = 2x112=224 T / hr Steam flow. Design Temperature = 540oC

Steam temperature after desuperheater = 200oC (Max. desuperheater).

Steam pressure at upstream of by = 6 ata (Max).

Pass valve

PAGE 105

TURBINE GOVERNING SYSTEM

10.1 INTRODUCTION

Power Station Turbines are constant speed machines. In our country these are

supposed to rotate always at a speed of 3000 revolution per minute (within a small

band of fluctuations on either side) to enable the coupled generator to produce

electricity at 50 Hz frequency.

The main purpose of governor is to maintain this desired speed of turbine

during fluctuations of load on the generator by varying steam Input to the

turbine.

The governing system in addition to ensuring the failing load-speed

characteristics of the turbine (i.e. a characteristic of failing output powerwith

raising shaftspeed above nominal value) also ensures the following functions:

i) The run up of the turbine from rest to rated speed and synchronising with the

grid.

ii) Meeting the system load variations in a predetermined manner, when

running in parallel with other machines.

iii) Protecting the machine by reducing the load or shutting off completely in

abnormal and emergency situations.

The governing system also includes other devices to protect the turbine from

abnormal conditions that may arise during operation.

10.2 METHODS OF GOVERNING



Basically there are three methods of varying the steam admission which are

briefly explained below (Fig. 1.12).

10.2.1 Throttle Governing

Here the supply of steam to the turbine is controlled through single batch of

nozzles either by a single valve or two or more valves operating in parallel. On

speed increase due to reduction in load on the machine, the throttle valve is

partially closed and as a result steam flow to turbine is reduced and the power

developed by the turbine is regulated.

10.2.2 By-pass Governing

In this system, in general, the steam is supplied through a primary valve and is

adequate to meet a major fraction of the maximum load which is called

economic load. At loads less than this, the regulation is done by throwing

steam through this valve. When the load on the turbine exceeds this economic

load which can be developed by the unthrottled, fuliflow through the primary

valve, a secondary valve, is opened and throttled steam is supplied downstream,

bypassing the first stage and some high pressure stages. This steam joins the

partially spent steam admitted through the primary valve, developing additional

blade torque to meet the increased load.

10.2.3 Nozzle Control Governing

Here the first stages are divided into number of groups, from three (3) in a

simple system, to six (6) or more in more elaborate arrangement. The steam

supply to each group of nozzles is controlled by a valve and the

numberofvalvesopenedisvariedaccordingtotheloadontheturbine.insuchcases,ifsa

y,seven(7)valves are opened to meet any given load condition, then six of them

remain full open and the actual regulation will be done by modulating the

seventh valve.

PAGE 106

10.3 SPEED SEN SING DEVICE

As mentioned, the job of governor is to vary the steam admission according to

variation of it is imperative that some form of s~ sensors (also ca#ed speed

governors) are required. Speed sensors sense the changes in magnitude of s~ from

the desired value and generate corresponding correcting ~s to control steam flow.



(Fig. 1 0.1 (a) shows a typical speed governing loop in block diagram form.

The system is in s~ equilibrium till the turbine torque T, is equal to the generator

torque T,. If there is a sudden fall in load demand, the excess torque developed by

turbine AT will accelerate the machine ad a rate.

α= ΔT/1

Where 1 is the machine inertia. As a machine s~ rises, the speed governor acting

through the control system will throttle the steam valves unite the turbine torque is

equal to the new power demand. (Fig. 1 0.1 (b) shows the response of turbine to a

sudden change in load and the resultant change in s~.

The percentage charbge in rated s~ corresponding to 100% change in load is

termed the "speed regulation" or the mdroopm of the turbine. It is nomially

around 4%. The regulations of the turbines operating in parallel Influence the load

sharing amongst them.

FIG. 10.1 (B)



PAGE 107

10.4 TYPES OF SPEED SENSORS 1 GOVERNORS

10.4.1 Mechanical

The earliest known automatic turbine speed governor was a mechanical

governor of centrifugal fly ball type, the governor being driven by the turbine

main shaft directly or through gearing. The centrifugal force acting on two

revolving weights, is opposed by the elastic force of a spring, so that the weight

take up different radius for each speed and produce a proportional displacement

of the sleeve linked to the fly balls through hinges (Fig. 10.2).

FIG. 10.2 FLYWEIGHT GOVERNOR BALANCES THE FORCE

OF SPRING AND WEIGHTS, MOVES SPEEDER ROD

10.4.2 Electrical

The electrical governor is a more recent innovation and made practicable by the

development of robust servomechanism and circuit components. An AC

generator, driven by the turbine shaft provides an electrical signal of a frequency

proportional to the speed. A frequency sensitive circuit produces voltage

proportional to this frequency. This voltage, after amplification, is fed to a

torque motor which in turn produces a proportional displacement.



10.4.3 Hydraulic

In simple form, a hydraulic governor for a turbine consists of a centrifugal

pump driven from the turbine main shaft. The pressurised oil from it being fed

into a cylinder containing a spring loaded piston. The oil pressure is

proportional to the square of the speed so the position of the piston also

becomes a function of the speed.

10.4.4 Hydro-Mechanical (Used in 21 0 MW BHEL (LMW) Turbine)

Here speed transducer is usually mechanical centrifugal type speed governor,

controlling through a combination of hy drautic relays & linkages. Oil for

hydraulic system is supplied by the main oil pump, which may supply oil to

lubricating oil system also at a reduced pressure.

PAGE 108 10.4.5 Electro Hydraulic (Used in 21 0 MW 500 MW BHEL (KWU) Turbines, in

parallel with hydraulic governing)

Due to large interconnected systems and growing automation of turbine

generator sets, governing system has to meet many additional requirements. The

combined advantages of electrical measuring and signal processing (flexibility,

dynamic quality, and simple representation of complicated functional

relationships) and hydraulic controls (continuous control of large positioning

forces) provide a very good combination.

The important characteristics of electro-hydraulic governing are (a) exact steady

state regulation with high sensitivity (b) safe load shedding by avoiding any

speeding up along the steady state regulation characteristic (c) possibility to

adjust steady state regulation in fine steps.

In electro hydraulic governing all transducers are electrical / electronic

components. The acquired signals (of control valve lift, speed, load & initial

pressure etc.) are processed electronically and processed signal is introduced at a

suitable point in the hydraulic circuit through a electro-hydraulic converter which

is used as a connecting link between the electronic modules and valve actuators.

Hydraulic signal before application to control valves servomotors is suitably

amplified.



Change over from electro-hydraulic governing to hydraulic governing is possible

during operation. When one system fails, other one comes into operation

automatically.

The displacement of the piston in a hydraulic governor, of the torque motor in

the electrical governor, or of the sleeve of the flyball governor, which is a

function of the shaft speed, is used to control the throttle valve of the turbine

through hydraulic relay action or mechanical linkages or a combination of both.

The arrangement of the various devices varies for different types and makes of

the turbines but their basic functions are similar.

10.5 MAJOR COMPONENTS OF GOVERNING SYSTEM

1) Governing Devices

a) Speed governor with pilot valve

b) 'Speeder Gear or Load-Speed Changer

c) Load Limiting Gear and

2) Protection Devices

a) Emergency trip valve

b) Over speed governor

c) Acceleration governor

d) Overspeed Limiting gear

e) Pre-emergency governor Low vacuum run back / unloading unit

g) Initial pressure regulator/ low initial pressure unloading unit (Gear).

10.6 BRIEF DESCRIPTION OF GOVERNING & PROTECTION SYSTEM &

THE DEVICES

The stop valves & control valves in the steam lines to the turbine are actuated by

hydraulic servomotors. The servomotors consist of a cylinder and a spring loaded

piston which is held in open position by admission of high pressure oil, against the

spring force, which ensures positive closing on the oil being drained out. The high

pressure oil supplied by the oil pump to the governing system is fed to the

servomotors through their pilot valves. The position of the pilot valve determines

the opening or closing of the servomotor. The high pressure oil which actuates the

servomotor is usually termed as "Power Oil" or "Sensing Oil".



The pilot valves of the stop valve servomotors are positioned in "OPEN" position

by yet another branch of oil called "Protection Oil / "Trip Oil" either directly or

through hydraulic relays. This "Protection Oil" is the same high pressure oil

butsuppliedthrough an EmergencyTripValve". This EmergencyTrip Valve, in

"Reset" position, admits oil through it, to be supplied to the various pilot valves of

servomotors, thereby enabling the opening of the stop and control valves. In

"Trip" position it shuits off the oil supply and drains out the oil from the lines

downstream of it, there by ensuring the quick closure of the stop & control valves.

PAGE 109

The protection oil being fed to the HP control valve servomotor piloit actuating

device is regulated through the Speed Governor Pilot Valve.The change in speed

which causes a corresponding change in the governor pilot valve, varies the oil

pressure which in turn regulates the position of the control valve through the

servometer and its pilot valve. There are variations in this arrangement. The other

governing and protection devices like load limiting Gear, LowVac. Pay off unit, 1

nitiai Pressure Regulator are hooked up into this control valve governing system

through hydraulic relays and linkages. Two different arrangements of control

system of HP control valve are shown in (Fig. 10.3 (a) and (b).

10.7 GOVERNING OF REHEAT TURBINES

In reheat turbines in cases of partial or full load thro woff even after the HP

control valves are fully closed, the entrained steam in the reheaters and hot reheat

line is more than enough to speed up the turbine above over speed limits. Hence

itis necessary to provide stop valves and intercoptor valves on hot reheat line

before IP turbine. While the stop valve is operated controlled similar to HP

control valve but at a higher speed range by a secondary or pre-emergency

governor as it is called. The valve remains full open at rated speed and starts

closing at about 3% over speed and is fully closed at about 5% over speed.

10.8 GOVERNING DEVICES

10.8.1 Speed Governor Pilot Valve

The pilot valve consists of a movable sleeve with ports for oil inlet, outlet and

drain inside which the double bobbin valve moves, actuated by the centrifugal

flyball governor. The high pressure oil or protection oil is admitted through the



inlet ports of the cylinder and sleeve. The pressure of the outlet oil to the control

valve servomotor pilot valve actuating relay is regulated by the relative

displacement 1 position of the piston and sleeve. Any change in shaft speed

produces a corresponding change in the pilot valve position, thereby causing a

corresponding change in the control valve opening.

10.8.2 Speeder Gear

A speeder gear is essential in the governing system of turbine of electricity

generating stations, for synchronising the machine with the grid and to vary load

when operating in parallel.

Speeder gear is needed to match the speed of the turbine to that of grid while

synchronising. After synchronising, the speed being determined by the grid

frequency, the speeder gear is used to raise or lower the load on the machine. It is

explained earlier how the relative position of the piston and ported sleeve of the

governor pilot valve regulates the control oil. While the piston is actuated by the

governor, the sleeve is operated by the speed gear. Hence at a particular position

of the piston movement of the sleeve varies the oil pressure. The speeder gear is

either operated manually from local or by a small motor, from remote.

10.8.3 Load Limiting Gear (LLG)

This device is incorporated in the governing system to limit the maximum opening

of the HP Control Valves to the desired upper limit. This may be done

mechanically by stopping movement of linkages connected with relays in the

control system or by limiting the sensitive oil pressure in the hydraulic system

thereby restricting the movement of speed relay by shutting off or draining the oil

as in the case of a system shown in (Fig. 10.3 (a) and by acting similar to a relief

valve, not allowing the control oil pressure to raise more than the preset value in a

system as shown in (Fig. 10.3 (b). It is to be noted that LLG provides only an

upper limit. A small motor is provided usually for remote operation of the gear.



PAGE 110

PAGE 111

PAGE 112

10.9 PROTECTION DEVICES

10.9.1 Emergency Trip Valve

The function of the valve has already been explained. (Fig. 10.4) shows a typical

trip valve in "Reset" position with provision for manual and oil injection tripping



and resetting. Remote tripping by Emergency Push Button also can be

incorporated to trip the trip valve position through a solenoid.

10.9.2 Over Speed Limiting Gear (OLG)

The purpose of the OLG is to limit the overspeed which would occur, should a

sudden loss of load take place. The control operates on each of the Emergency

Stop Valves. On RH turbines, it operates on both the HP and IP Emergency Stop

valves. The gear comprises an additional solenoid operated pilot valve which

releases oil from the steam stop valve power cylinder, when the solenoid is

energised, valve closes rapidly under the action of the spring. There are two sets

of contacts in series. One is operated by a Wattmetric relay and the other by a

relay under steam pressure in a selected range of the turbine. For example, the

contact operated by steam pressure may be set to remain closed over the range

60% to 1 00% load and that operated by the load at Oto 30% load. If a load

exceeding 60% is suddenly reduced to less than 30%, the load operated contact

will close at once but the pressure operated contact will not open immediately

because steam already in the turbine continues to expand thus with both contact

closed the solenoid is energised and the emergency stop valves close. The

solenoid will de-energies and open the valve when the Steam pressure drops and

the contact opens.

10.9.3 Acceleration Sensing Device

To cope with the rapid rise in speed that would occur in the event of a sudden

loss of load, san acceleration sensitive governor is fitted. This causes the rapid

closure of HP throttle and IP interceptor valves and when the acceleration of

theturbine cases the speedwould have come down generally to such avaluethatthe

main governor will keep these valves closed.

If for any reason the speed should continue to rise the over sp;eed governor will

come into action and trip the turbine causing closure of the HP and IP emergency

stop valves.

A typical acceleration governor consists of two concentric tubes which normally

rotate together. The inner tube is driven through gearing by the turbine rotor.

The outer tube carries an inertia wheel and is driven by the inner tube through a

torsional spring. There are ports in each tube connected to the HP control oil and



IP control oil circuit. Under normal conditions these ports remain closed.

During high acceleration periods, extra spring force is required to accelerate the

outer tube and inertia wheel, resulting in the inner driving tube advancing its

position relative to the outer tube. This movement aligns the ports in the inner

and outer tube and draining the control oil with resultant rapid closure of the

valves. (Fig. 10.5)

10.9.4 Pre-Emergency Governor

The IP control valves / Interceptor valves independently or along with the HP

control valves are operatged either by another centrifugal governor or the same

main speed governor by an acceleration sensing differentiator and pilot valve. The

operation is similar to that of HP control valve governing. The interceptor valves

are normally full open at rated speed. in a typical case, they would begin to close

at a speed about 3% above normal and completely close at about 5% above

normal. The final speed rise due to time delay in relay operation and the effect of

steam in loop pipes, is about 7 to 8% above normal. The IP or Pre-emergency

governor is normally present and is not changed during normal operation.

PAGE 113

PAGE 114



10.9.5 Over Speed Governor 1 Emergency Governor

ln the un likely event of speed increasing to lll% to ll2% emergency or over speed

governors are provided to trip the machine. The emergency governor is normally

mounted on the front end of HP rotor. It consists of two strikers for reliability.

Centre of gravity of the striker is away from the centre line of rotation and they are

held in place by springs. As soon as the speed rises to lll% -112%of the nominal,

centrifugal forceon the strikers overcomes force of the springs and the strikers fly

out: once the strikers begin to move out, distance between the centre of gravity of

the strikers and centre line of rotation increases, causing further increase in

centrifugal force. Thus strikers, once dislodged from stable position would

continue to move out till checked by the stop of the body. One or more of the

strikers flying out will hit trip levers placed close the the tip which on being hit

will instantaneously trip emergency trip valve causing the complete shut off of all

steam input to the machine (both HP and I P).

10.9.6 Testing of Strikers by Oil Injection

It is necessary to test these strikers for their free movement periodically. This is

done by oil injection, without actually raising the speed of the machine. Actual

over speed test can be done in some turbines during start up or shutting down by

tripping the machine. Provisions are also made in some turbines to test during

operation when the Emergency trip valve is kept gagged in "Reset" position or the

over speed trip level made out of action during the test period. A tell tale

indication is normally provided to show that the striker has flown out.

10.9.7 Low Vacuum Unloading Gear



In addition to normal control of the turbine a Low Vacuum unloading gear may

also be provided. The purpose is to gradually close the turbine CVs when the

pressure in condenser rises above a preset value. This is achieved by different

arrangements. The displacement of a bellow connected to the condenser vacuum

space is utilised to reduce the control oil pressure directly or by changing the

position of pilot valve thus reducing the control valve opening or closing a contact

which will undown the LIG motor.

If required, a provision is made on the device by means of which automatic

reloading of the turbine with restoration of condenser vacuum is lirevented and

reloading will be manual. The stroke of the unloading gear is so limited that ft can

only close the valves down to no load position. This is to prevent the machine

from motoring (Fig. 10.6).

PAGE 115

PAGE 116

10.9.8 Initial Pressure Regulator (IPR)

The device is similar to low vacuum unloading gear. If the throttle pressure

should fali to more than 10% below normal, the device comes into operation and

starts closing the CVs until a balance is reached. It is provided with a stop to

avoid motoring and also equipped with a cutout device to block the regulator



completely to permit starting with lower boiler pressure. Reloading is automatic

with restoration of throttle pressure. If required it is provided with a manual

resetting device to prevent automatic reloading of the machine.

10.9.9 On Load Testing Device

The risk of serious damage to a large turbine due to sticking of one of the steam

valve in the event of sudden loss of load is so great that provision is normally

made for testing at suitable intervals of time, the freeness of movement of each

valve steam while the turbine is running on load. In general, this is done by a

device which releases the power oil from under the servomotor / relay piston of

the valve being tested. Provision is made such that only one set of valve can be

tested at a time with machine load limited to the capability of the other set.

10.10 GOVERNING & PROTECTION SYSTEM OF 210 MW (LMW) TURBINE

10.10.1 Introduction

High response hydro-mechanical governing system has been for the steam turbine

to maintain the speed at the desired set points during startup and normal operation.

It also serves to prevent undesired overspeeds following sudden loss of export

load. The control action is of the proportional type with a steady state overall

speed regulation (i.e. proportional band) of 4 ± 1 %. This proportional band is

necessary in order to realise:

a) Stable speed control in isolated operation of the set: and

b) the desired degree of load distribution between sets running in parallel.

All the operations of starting and loading of the set can be performed by

manually operating the speeder gear hand wheel located at front standard or by

operating speeder gear motor remotely from unit control panel.

10.10.2 Special Features

a) In the event of generator breaker opening, following a full load loss, governing

system prevents the over speeding of the set to a dangerous level and quickly

stabilises the set on house load or on no load. This feature enables the quick

reloading in case of surplus loss of export load. This is achieved by electro-

hydraulic transducer which closes the control valves when generator circuit-

breaker opens. However, the electro hydraulic transducer receives the signal

only for two seconds.



b) The governing system envisages anticipatory over speed control gear termed as

"Differentiator". The differentiator causes anticipatory closure of the turbine

control valves depending on the magnitude of the acceleration being

experienced by the turbine. This action prevents the transient speed rise to

dangerously high levels.

c) Transient speed rise is anticipated to be not more than 7 to 8% over the normal

speed, even in case of total loss of export load.

d) Speed governor can control speed in the range of approx. 300- 345O rpm

when set is not synchronised.

e) Load limiter has been foreseen to avoid accidental overloading of the set. The

set point of load limiter can be chosen over entire range from no load to full

load.

f) Governing gear operates on constant oil pressure principle.

g) Lnitial steam pressure unloading gear (ISPUG) has been foreseen to unload the

set in case initial steam pressure drops by more than 1 0% of the rated value.

PAGE 117 10.10.3 BRIEF DESCRIPTION OF GOVERNING REGULATION SYSTEM

10.10.3.1The governing or regulation process is achieved by a combination of

mechanical and / or hydraulic signals causing interaction of elements listed

below:

SI.NO. Description Tag No.

a) Speed governor 01

b) Follow-pilot valve 02

c) Summation-pilot valve 03

d) I ntermediate-pilot valve 04

e) Control valves servomotor 05

f) Load 1 speed control pilot valve 06

g) Load Limiter 07

h) Differentiator (Anticipatory gear) 08

i) Electro-hydraulic transducer (E.H.T) 09

j) Speedergear 10

k) Initial steam pressure unloading gear (ISPUG) 11



l) Lever 12

10.10.3.2 The block diagram of the regulation process is given in (Fig. 10.7). The salient

features are explained below:

Governing 1 Regulation Process (Fig. 10.7) & (10.8)

a) Any difference between the power generated by the turbine and the load on

it, would cause change in the speed of the set. This change is sensed by speed

governor (01). The governor sleeve moves proportional to speed change.

b) The pilot spool of follow-pilot valve (02) follows the governor sleeve. This

movement is caused hydraulically and there is no mechanical linkage between

the governor sleeve and the pilot spool.

c) The follow pilot spool actuates the summation pilot valve(O3)through lever

(12).The summation pilot valve converts the mechanical signal into hydraulic

signal which actuates the intermediate pilot valve(O4). The intermediate pilot

valve also receives hydraulic signals from:

i) Differentiator (08)

ii) Electro-hydraulic transducer (09)

iii) Protection system

d) The intermediate pilot valve amplifies the hydraulic impulses.

e) The amplified hydraulic signal from intermediate pilot valve actuates the

control valves servomotor (05).

f) The control valves servomotor actuates the control valves of HPT and iPT

through rack, pinion, cams and other linkages.

10.10.3.3 Operation of Differentiator (Anticipatory Gear)

The purpose of this device is to prevent transient speed rise to dangerously high

level due to sudden rejection of a major chunk of load. Ditferentiator receives

hydraulic impulse from summation pilot valve and it senses the acceleration of

turbine. The differentiator cuts in only if the acceleration of the set is equal to or

more than the value corresponding to load dump of more than 50% of the rated

load.

The cutting in of the differentiator results in anticipatory closure of control valves

through intermediate pilot valve (04) and control valves servomotor (05).



PAGE 118

10.10.3.4 Operation of Electro-Hydraulic Transducer

Electro-hydraulic transducer cuts in, when electrical signal due to generator circuit

breaker opening is given to energise the operating coil of the transducer. This

would lead to anticipatory closure of control valves. The electrical signal tothe

electro-hydraulic transducer is applied onlyfortwo seconds so thatthis anticipatory

action exists onlyfortwo seconds. This has been done with the objective to

enablethe normal governing system to bring the set to no load position and not to

shut down the set due to dangerous overspeed.

10.10.3.5 Operation of Load Limiter

The load-limiter physically checks the movement of summation pilot spool in load

increase direction. As soon as the summation pilot spool presses the limiter rod,

an annunciation signal "Reduce Load" would flash in unit control panel. A

continuous operation of set on load at which load limiter is set, is not

recommended.

10.10.3.6 Operation of Initial Steam Pressure Unloading Gear

Initial steam pressure unloading gear (ISPUG) starts unloading the set in case

steam pressure ahead of emergency stop valves fails below 90% + 2% of the rated

valve i.e. below 1 17 + 2 ata. The actuation of, ISPUG causes a pressure drop in

secondary sensitive oil line and thus closing the control valves. ISPUG would

completely unload the set if initial steam pressure falls to 91 ata and maintains the

set at no load condition. In between 1 1 7 + 2 ata and 91 ata unloading is

proportional to pressure drop, In case steam pressure is restored, it reloads the set.

10.10.4 BRIEF DESCRIPTION OF PROTECTION SYSTEM

10.10.4.1 Protection system functions similarto governing system by a combination of

mechanical and hydraulic signals on various elements described below:

SI. No. Description Tag No.

a) Emergency governor 21,

b) Emergencyh governor levers 22

c) Emergency governor indicators 23

d) Emergency governor testing cock 24

e) Emergency governor pilot valve 25



f) Emergency stop valve servomotor 26

g) Interceptor valve servomotor 27

h) Turbine shut down switch 28

10.10.4.2 Protection system operates and trips the set (complete closure of emergency

stop valves, interceptor valves and control valves) in the unlikely event of the

following hazards.

a. Speed rise upto 1 1 1 to 1 1 2%

b. Speed rise upto 114 to 11 5%

c. Thrust pad wear beyond impermissible value

d. Vacuum below 540 mm Hgcol.

e. Lube oil pressure below 0.3 atg. Main steam temperature low.

f. H.P. heater level extra high.

PAGE 121 10.10.4.3 Manual tripping of the set may be done by pressing the knob of shut down

switch.

10.10.4.4 Automatic closure of emergency stop valves occurs,d control oil pressure

drops below 10 atg.

10.10.4.5 Automatic closufre of interceptor valves occurs, if control oil drops below 5

atg.

10.10.4.6 Protection Against Overspeeding By 11% to 12% Above Nominal

Speed

a) Emergency governor strikers, which are eccentric to the axis of rotation,

flyout against spring force at a speed 1 1 % to 12% more than rated speed.

b) Stikers strike one end of the emergency governor levers. The other end of

lever presses impulse pilot spools downwards.

c) The impulse pilot spool converts the mechanical signal into hydraulic

signal and actuates the emergency governor pilot valve to trip the set.

The emergency governor pilot valve receives hydraulic signal from

follow pilot valve when the speed of the set rises by 14% to 15% above rated

speed. The emergency governor pilot valve receives hydraulic impulse from

the turbine shut down switch also. Emergency governor pilot valve, in turn,



actuates to close the servomotors for the emergency stop valves, interceptor

valves and control valves.

10.10.4.7 Protection Against Over-Speeding by 14% to 15% Above Nominal

Speed

a) In case protection system for 1 1 % to 12% speed rise fails to trip the set,

speed may go even higher to dangerous vlalves. For this purpose, back up

protection to trip the setata speed l4%tolS%above nominal is provided.

b) At speed 14% to 15% above rated, follow pilot spool moves towards

extreme right position and gives hydraulic signal to emergency governor

pilot valve (25).

10.10.4.8 Manual / Automatic Tripping By Turbine Shut Down Switch (28)

By pressing the knob or by energizing the solenoid, a hydraulic impulse to

emergency governor pilot valve is transferred. EGP valves, in turn, actuates to

close the servomotors of emergency stop valves, interceptor valves and control

valves.

10.11 210 MW / 500 MW (KWU) TURBINE GOVERNING SYSTEM (FIG.

10.9)

10.11.1 Introduction

The speed of turbine generator set can be controlled either by hydraulic governing

system or by electro hydraulic governing system. Changeover from one governing

system to the other is possible during operation. When electro-hydraulic

governing is controlling the speed of the set, hydraulic speed governing acts as

back up and comes into operation automatically in case the former fails.

10.11.2 HYDRAULIC SPEED CONTROL

On the shaft of main oil pump, the "Hydraulic Speed Transmitter" has been

provided, which provides primary oil pressure. The change in absolute pressure

can be taken as proportional to small changes in speed (within limits of steady

state characteristics). This primary oil pressure acts on diaphragm of "Hydraulic

speed governor" against the force of speed setting spring, which is compressed by

speed changer. Travel of diaphragm is limited by start up and load limiting

device. The movement of diaphragm is transmitted by link mechanism to



"Auxilliary Follow up Pistons". The piston of "Auxilliary Follow up Pistons" is

held in balance by a spring against oil pressure.

PAGE 123

The oil is essentially trip oil fed from trip circuit and drained from a port formed

between the piston and the sleeve. Depending upon Port opening, oil pressure gets

stabilised corresponding to initial displacement of the piston and initial spring

tension. The necessary auxilliary oil pressure provides asignal to "Hydraulic

amplifier" through its pilot.

The piston of the hydraulic amplifier assumes a position corresponding to

secondary auxilliary oil pressure and actuates the sleeve of follow up pistons,

through a linkage system. A feed back system has been foreseen for quickly

stabilising the position of the pilot valve and the piston of "Hydraulic Amplifier".

With a view to avoiding stricking of pilot spool and increase sensitivity, pilot

spool is kept rotating due to reaction of oil leakage through tangential holes by

means of control oil.

Asecondary oil pressure corresponding to the position of the sleeve and related

spring tensions is built up in the "Follow-up Pistons" of thej'Hydraulic Amplifier"

in a similar fashion to "Auxilliary Follow-u Pistons'. Any change in the position

of,Iink results in proportional change of secondary oil pressure in the "Follow-up-

Pistons" of "Hydraulic Amplifier". The secondary oil circuit is also fed oil from

trip oil circuit through reducing valves.

The varying secondary oil pressure'in the "Follow-up Pistons"..' of "Hydraulic

Amplifier: operates the control valves.

10.11.3 Electro-HydraulicControl

(Fig.10.10) is a typical principle diagram of Electro-Hydraulic Governing. The

turbine speed control loop employs electronic means of speed measurement. This

provides a ,faster,. speed of response than the mechanical method.

The unit power output control present in the scheme deals with the non linearities

of the Governor illustrated in (Fig. 1 0. 1 1). The, desired value setting of power

controllers forms the means of establishing the generated output at a particular

frequency value. Its operation must of course be relatively slower than that of the



governor control loop in@ order that the latter may be effective in dealing with

emergency conditions.

With the ability to apply an adjustab!e load / frequency characteristic to each unit

of a part of the grid system it is possible to vary the overall response of the units to

frequency change and also vary the distribution of load changes thus involved

between the units.

The turbine speed is measured digitally by an Electrical speed transducer mounted

at the HP end of turbine rotor. An electrical controller consisting of electro-

hydraulic converter and a moving coil system link electrical and hydraulic parts of

governing system. A control sleeve is coupled to the moving coil (position of

which is controlled by electric controller)., The control sleeve determines the

position of pilot ,valve and this determines the secondary oil pressure th 11 rough

"Follow-up-Pistons".

10.11.4 Changeover from Hydraulic to Electro-Hydrautic Control

As earlier pointed out, changeover from one control system to the other one is

possible during normal operation of turbine, since the two system are brought into

parallel connection after associated follow up pistons which represent a minimum

value selection, meaning that the system with the lower reference value is always

the controlling one.

If the turbine is to be run under hydraulic control,, the reference speed set point of

the electrical controller is t "Max. Speed", which prevents the efectro-hydrauiic

system from coming into action. When bringing in, the efectro-hydraulic control

system, the reference speed set point of the electrical controller should be reduced

slowly until the secondary oil pressure drops slightly. When this occurs, the

electrohydraulic converter has taken over. The speed changer of the hydraulic

speed governor is then set at maximum speed. The electro-hydrautic converter is

now fully effective and can operate over the entire output range. The hydraulic

speed governor also acts as a speed limiter in the event of electrical controller

developing a fault. In this case, operation of turbine may immediately be

continued by means of hydraulic speed governor.

PAGE 124



10.11.5 Changeover from Electro Hydraulic to Hydraulic Control

PAGE 125



FIG. 10.11 LOAD LOOP CHARACTERISTICS OF GOVERNOR PAGE 126

10.11.5 Changeover from Electro Hydraulic to Hydraulic Control

This changeover is done in the reverse sequence mentioned above. First, speed

change is actuated in the 'decrease" direction until secondary oil pressure drops

slightly. This is an indication that hydraulic speed governor has taken over the

control. Then reference speed limit of electric controller is set at ..maximum".

Now the hydraulic speed governor is fully effective, and can operate over the

entire load range.

The secondary oil pressure is transmitted to the actuators of the HP and IP

control and controls their opening.

10.11.6 Advantages

The integrated electronic and hydraulic control systems offer significant

advantages. These are given below:

a) Exact load frequency drop with high sensitivity.

b) Reliable operation in case of isolated power grids.

c) Dependable control during load rejection.

d) Low transient and low steady state speed deviations under all operational

conditions.

e) Excellent operational reliability and dependability.

Safe operation of the Turbo set in conjunction with the turbine stress

evaluation (TSE).

The other features are:

i) A sequence timing device which adjusts the relative opening of HP and I P

control valves and thus avoids heating of HP exhaust at reduced loads.

ii) Two load shedding relays which act for anticipatory closure of control

valves in the even to flarge load dump.

iii) An extraction valve relay which actuates NRVs in extraction lines as

demanded by load situation.

10.11.7 Protection System



The protection system has been designed to protect the turboset from any mishaps

by closing the fast acting stop and control valves and thus tripping the set. The

hydraulic trips are:

a) Overspeed trips 1 and 2

b) Low vacuum trip

c) Thrust bearing trip

d) Local manual trip

The Electrical trips are:

i) Manual remote trip

ii) Low vacuum trip

iii) Low lube oil trip

iv) Fire Protection trip

v) Trips due to other causes e.g. generator protection.

The electrical trips act through remote trip solenoids for tripping the set.

All the protections act for closing of EVSS, IVs, HPCVs and IPCVs through

Main trip valves.

10.11.8 Automatic Turbine Tester

The turbine operates uninterruptedly for long slots of time and none of the

protection device may be operating for years at a stretch. During this healthy

period of turbine operation, it is necessary to check the availability and efficiency

of the protection devices. The scheme envisages an automatic turbine tester which

carries out the functional tests of various devices even when turbine is carrying

load without PAGE 127

jeopardizing the safety of the turbine.

Automatic Turbine Tester consists Of two sub groups, one sub group is for safety

devices e.g. Remote trip device, overspeed trip, thrust bearing trip and hydraulic

low vacuum trip. The other sub-group is for in testing main stop valves, main

control valves, reheat stop valves and reheat control valves. Testing of ic valves is

done at reduced load.

10.11.9 L.P. Bypass System



An additional control to dump the excess steam of the reheater in the event of

violent load excurison is desirable. This is achieved by incorporating a 1 00% LP

Bypass System. The LP System when operating in conjunction with HP Bypass

keeps in matching of steam parameter during start-up. The control of L.P. Bypass

are integrated with EHG controls. It dumps the steam from the reheater to the

condenser in case the reheater pressure is more than the desired value being

demanded by the load considerations.

10.11.10 Special Features

i. The system is equipped with the electro hydraulic control of gland steam

pressure which maintains at all gland bleed points a uniform pressure as

determined by preset value.

ii. Damping Devices are envisaged in the control oil circuit (secondary oil) to

damp out any pulsation which may ocur in the flow to the control valves.

Thus it ensure the smooth operation of the system.

iii. The system is provided On-Load testing facility for all extraction fine non-

return valve.

iv. The emergency stop valve HP control valve, Interceptor valve IP control

valve and LP Bypass stop;

valve and control valve are housed in their single combined casing'

v The pilot valves in the control circuits have rotational and oscillatory

motion.

Through rotation, minimum friction is created between pilot valve and sleeve.

Oscillation prevents seizing of pilot valve in the guides and other sliding surface.

vi LP Bypass cuts into operation when the water is charged and vacuum is

healthy.

PAGE 131

AUTOMATIC TURBINE RUN UP SYSTEM

11.1 INTRODUCTION

With increase in unit capacity associated with increased capital costs and the steep

rise in fuel cost, it necessary to maintain the availability of thermal power sets at

as high a level as possible. To achieve th it is, essential to reduce the extent of



damage, to provide facility for tracking down the faults and th causes by m good

an over view of the process, and shortest possible start up and shut downs.

For start up, acquisition and analysis of a wide variety of information pertaining to

various paramete of steam turbine demand, quick decisions and numerous

operations from the operating personnel. orderto reducethe arduoustask of

monitoring various parameters and effect sequential start up, minimi the possible

human errors and to achieve start up in minimum time in optimum way Automatic

Turbi Run up System (A.T.R.S.) is introduced.

11.2 PHILOSOPHY

The ARTS is based on functional group philosophy i.e. the main plant is divided

into clearly defin sections called functional groups such as oil system, vacuum

system, turbine system. (Fig. 1 1.1). Ea functional group is organised and arranged

in sub group control (SGC), sub loop control (SLC) and contr interface (Cl). Each

functional group continues to function automatically all the time demanding enab

criteria based on process requirements and from neighbouring functional groups if

required. In t absence of desired criteria, the system will act in such a manner as

to ensure the safety of the ma equipment.



PAGE 132

11.3 FUNCTION OF THE SUB-GROUP CONTROL (OIL)

The main task of the sub-group control "Oil" during a start up programme are

(a) establishing lubricating oil, control oil, and jacking oil supply

(b) putting the turbine on barring gear

(c) taking off the jacking oil pump when-appropriate turbine speed is

achieved

(d) taking off the AOP when turbine main oil pump has taken over

(e) leaving the various sub loop control of oil systems in automatic regime so

that the logics built up in this sub-loop control will take care of the

requirement of the individual systems.

The sub loop control of any system essentially serves to provide an auto

start / stop command when certain preset conditions are fulfilled. The feature

can be included or excluded by manual push botton commond or by sub-group



control. During a shut down programme this sub group control essentially

serves to bring the turbine to a standstill and switch off the oil system.

11.4 FUNCTION OF SGC TURBINE

The main tasks of the sub group control (SGC) "turbine" during a start up programme are

(a) Warming of the admission pipe lines and stop and control valves (carried

out by the warm up controller in conjunction with the turbine stress

evaluator)

(b) Warming up of the turbine at 64ORPM (carried out by electro hydraulic

speed controller, in conjunction with the turbine stress evaluation)

(c) Acceleration of the turbine to synchronous speed carried out by the

turbine stress evaluator,

(d) Synchronise the machine to the grid (carried out by the auto synchroniser)

and

(e) block loading of the machine.

In addition, the sub-group control also switches on the sub-loop control

drains during the course of its programme for Warming up of casing pipe

lines and other valve bodies.

During a shut down programme the sub-group control "Turbine" essentially

does the task of

(a) re-loading the machine and switching on sub-loop control drains

(b) Setting the load controller to its minimum value and lowering the speed

reference valve of the speed controller

(c) Tripping of the machine.



FIG. 11.2 OPERATOR INTERFACE TO SUBGROUP

PAGE 133 11.5 FUNCTION OF SGC CONDENSATE AND EVALUATION

The sub-group control for condensate and evacuation system accomplised its task

which comprises of keeping at least one of the two CEPs in operation / evacuating

the condensate gases from the system, maintaining the desired level of condenser

pressure when turbine in operation and breaking the vacuum as and when required

by Mechanical process.

The combination of the three sub-group control will provide a turbine start

up in the ollowing sequence :

1) Prepare turbine for start-up

2) Start oil supply system and turbine gear

3) Start up condensing plant

4) Start seal steam system

5) Warm up main steam line

6) Warm up turbine

7) Accelerate turbine to rated speed



8) Synchronize and block generator

FIG. 11.3 HIERARCHY OF CONTROL

11.6 SHUT-DOWN OPERATION

The shut-down is carried out in the following sequences :

f. Reduce turbine load

g. Ensure bypass operation

h. Reduce turbine load to less than 5 % of rated capacity

i. Switch off generator

j. Trip turbine and check if AOP is in operation

PAGE 134

k. Shut down steam generator

l. Close Main Steam Valve of Boiler

m. Depressurise main steam line

n. Close drains of main steam line

o. Shut down condensing plant

p. Start turning gear

11.7 SUB-GROUP CONTROL (S.G.C.) (FIG. 11.4)

A sub group control executes commands, to bring the equipment upto a particular

defined state and contains the start up and shutdown programme of there spective

group S.G.C. issues commands efther to control interface level on switches on

S.L.C. Desired number of criteria act as preconditions before the S.G.C. can take



off to execute its defined programme. The programme comprises of steps. For

each step there is provision form on during time and waiting time. Waiting time

implies that the subsequent step could not be executed unless the specified time

elapsed. The time between command signal and appearance of check back signal

is known as monitoring time and when it is executed alarm is initiated and the

programme is not proceeding further. In addition execution of any step is also

permitted only if the conditions specified for that particular step are completely

fulfilled. By-pass conditions are included to enable switching on after S.G.C. at

any stage after completing certain task manually d so desired.

FIG. 11.4 SUBGROUP CONTROL STRUCTURE

11.8 MODES OF OPERATION

ATRS can be operated in the following three modes:

a) Automatic : In this mode all the specified operations are

carried out in appropriate sequence automatically.

PAGE 135

b) Step Mode : This mode has been envisaged to allow the operator to by

pass anychte had the operator finds that a criteria is not being obtained

because of the, malfunctioning of any instrument or transducer.

c) Operator Guide Mode : in this mode the ATRS does not issue any

commands. The command outputs are blocked and have to be issued



only by hand. This mode allows correct fulfilment of the criteria and

satisfactory progress of the sub-group control system to be tested during

commissioning and at other times. This mode can also be used as a

training aid for the operators.

11.9 CONTROL INTERFACE (Cl)

The control interface module forms the link between the individual commands and

the power plant (Fig. 11.5 & 11.6). Each remote controlled drive has a control

interface module. The module consists of command section, monitoring section,

power supply and alarm section. The command section proves the control

actuation signals to the interposing relays in the switch gear. Soienoid valves can

be actuated directly up to certain capacity (36 W). The, monitoring section

normally checks the command functions namely the position of the drive check

back signals, protection logic and FGC.

The control interface module type AS 1 1 is used mainly for ON /OFF motor

drives. AS 12 for motor operated regulating valves, and AS 13 for solenoids. The

Cl modules monitors status discrepancies, running time between command output

and check back actuation of the torque switches, check back for non contincidence

+ 24 V supply voltage for module control circuit MCB in the switch gear and

blocking of command by protective logic. Protection commands are given

priority.



FIG. 11.5 DRIVE CHARGEOVER AUTOMATIC

PAGE 136

11.10 SUB LOOP CONTROL (SLC)

When SLC is switched ON, it actuates the connected mechanical equipment to the

required operating condition as per the process condition. No sequence logic is

involved. SLC can be switched ON / OFF manually or through SGC.

11.11 CONTROLS AND DISPLAYS

S.G.C. Switching ON/OFF can bedonefrom desktile (Fig. 11.2). Display of step

and criteriaare available at the control desk. PB-2 is for switching ON and OFF.

The programme can be executed manually at the Cl level when the SGC is OFF.

PB-1 is for start up and 3 for shut down in auto mode. Rapid flickering (8 Hz)

light in lamp 4 or 8 indicates that programme is running towards desired status,

"Steady" on completion of programme and also flashing light indicates that the

programme is in the desired mode but a programme fault has appeared.

FIG.11.6

PAGE 137

FIRE RESISTANT FLUID

12.1 INTRODUCTION

Mineral lubricating oils used in hydraulic systems situated adjacent to equipments

maintained at high temperatures have always constituted a considerable fire risk.



Fires have in fact occurred in the past in power stations, when oils and hydraulic

fluid leaking on to hot surfaces, have ignited. Even at ambient temperatures, fires

have occassionally resulted from such leaks owing to sparking from commutators

of adjacent electric motors. Since 1957, mineral oils used for hydraulic purposes,

have generally been replaced with fire resisting fluids. Typical power station

hydraulic system applications include circulating water valve operation, boiler

controls, boiler damper and soot blowing equipment and window operating gear.

Fire-resisting lubricating oils are now being introduced for use in air compressors

and also, for use in turbine control gearwhere the turbine lubrication system is

quite separate from the control system. Fire resistant fluid are used in the

governor control system (32 bar system) of 500 MW steam turbine of KWU

Design.

12.2 SPECIFICATION

This specification is valid for fire resistant fluids, hereafter referred to as FR F,

used as electro hydraulic control fluids in turbine governor systems.

12.2.1 Chemical Composition

Triaryphosphate esters are reaction products of phosphorous oxychloride and

phenol and phenol derivates, which are obtained either from natural raw materials

(Natural FR F) or synthetic raw materials ('synthetic'FRF). The final product must

be free of ortho-cresol-compounds. In order to improve certain properties, e.g.

corrosion protection, oxidation stability, additives may be admixed provided that

they have no negative effects on the materials of the FRF-system of its operation.

12.2.2 Requirements

a) CorrosionProtection:The FRF shall not cause corrosion to the following

materials: steel, copper, copper alloys, zinc, tin, aluminium. The FRF must

grant sufficient corrosion protection to the materials used in the FRF system.

The FRF will be continuously regenerated with a fuller's earth filtration

system.

The FRF must not cause any erosion or corrosion on the edges of the control

elements.

b) Viscosity Range : FRF of viscosity class ISO VG 32 and ISO VG 46 shall be

used



c) Shear Stability : The FRF must be shear-stable. It should not contain any

Vi-improver. (Viscosity index improvcfr).

d) FireResistance: FRF leaking from the system must not ignite or burn in

contact with hot surfaces (up to 550'c).

e) Thermal Stability: The FRF must be capable of with standing continuous

operating temperatures of 7500c without physical or chemical

degradation.

f) Compatability: The FRF must be miscible with traces of triaryphosphate

esters of another kind but of the same base ('natural'or'synthetic'). There

should be no deterioration of the FRF in the presence, ofsuch race quantities.

g) Physiological Consideration: The FRF must not represent a safety or health

hazard to the persons working with it providing that normal good industrial

hygience practices are following:

PAGE 138

12.2.3 Factors Affecting the Quality of Fire Resistant Fluids

During operation hydraulic fluide on the base of phosphate esters are undergoing

degradation in a more or less progressive manner. Degradation will be caused

mainly by OXIDATION AND HYDROLYSIS.

The loss of oxidation and hydraulic stability will be indicated by the

increase of the NEUTRALISATION NUMBER. Other indicators are :

- Increasing air release value

- Foaming and high from stability

- Increasing conductivity

a) Oxidation Stability : Oxidation stability will be affected by :

- High temperatures including local overheating

- Catalytic effect of dissolved metals

- High air content

b) Hydrolytic Stability : Hydrolytic Stability will be affected by :

- water content

- Presence of degradation products mainly of acidic nature which can

catalyse further hydrolysis.



c) Viscosity : The Viscosity can be altered by progressing oxidation or by

mixture of fluid and other non compatible liquids e.g, oil.

d) Auto-lgnition Temperature : The autoignition temp. will mainly be

affected by oil content in the control fluid.

e) Water content : Water content starts hydrolysis and must be avoided by all

means.

f) Neutralisation Number : Increasing neutralisation number indicates

oxidation and / or loss of hydraulic stability.

g) Air Release Properties : The air release properties are mainly affected by

materials non-compatible to the fluid among other oil is badly affecting the

air release properties. Air release properties must also be seen in

conjunction with the system.

h) Foaming : Factors affecting foaming can be :

- High air content

- Reduction of anti foam activities by filtering

- Contamination e.g. by oil.

i) Chlorine Content : Chlorine content starts acidic reaction. Chlorine finds

access if environmental conditions are not adequate. Also, cleaning liquids

can affect the fluid.

j) Purity : Purity can be affected by :

- Access of solid particles from external sources.

- Build-up of solid particles by degradation

k) Conductivity : Conductivity will be affected by :

- Degradation products

- Water

- Chloride

- Contamination from other sources

12.2.4 Fluid Testing Programme

a) Sampling Frequency : The effectiveness of the program is greatly

increased if a regular sampling interval is set up. Experience has shown the

following sampling frequency to be the most effective:

Turbine start up - Sample before and after start up



1st Month of operation - Weekly

1st Year of operation - Every month

2nd Year of operation - Every 2 months

3rd Year of operation - Every 4 months

Samples can be taken from the sampling value in the auxiliary filter system

Two separate samples are required due to the particle count testing.

PAGE 139

b) Test Limits :

Viscosity Minimum 200

Maximum 240

This is a test of the fluidity at 100 F and is important to pump life and system

response time. Experience has shown that fluid out of this range has probably been

contaminated with a mineral oil or fuel.

Mineral Oil Content 4% Maximum

Recordings of below 1% in this category are generally the results of hydrocarbon

extraction from pipe dopes, seals or packings contamination above 1% may

indicate contamination from an external source. A connection above 4% is

potentially harmful to the system due to the lowering of fire-resistance and

incompatibility with some materials of construction.

Water Content Maximum 0.15%

New Hydraulic fluid may contain as high as 0.10% moisture. A reading above

0.15% indicated that water is entering the system from an external source, such as

an intercooler. Corrective measures should be taken to remove the excess water

and to eliminate the source of the leak. High water content can lead to increased

pump wear and hydrolysis of FRF to form corrosive acids.

Neutralisation No. Maximum 0.50 MGKOH /G

An operational maximum of 0.30 MGKOH / G should be set on neutralisation

numbere with an understanding that 0.50 MGKOH / GM can be tolerated. Fullers

earth filtration may be used to control and reduce this value. A high neutralisation

number could increase valve erosion.

Chlorine Content Maximum 150PPM



A maximum of 150 PPM chlorine, by X-Ray emission method, can be tolerated in

operational systems. Higher chlorine values have been associated with servo-valve

erosion.

Particle Count Maximum Class 3 SAE A-6D

5 to 10 microns 24,000 particles / 100 ML.

10 to 25 microns 5,360 particles / 100 ML.

25 to 50 microns 780 particles / 100 ML.

50 to 100 microns 110 particles / 100 ML.

Over 100 microns 11 particles / 100 ML.

Fluid cleanliness, in terms of particle distribution, has been shown to be of

utmost importance when operation servo-valves.

Resistivity Minimum 5.0 x 109 OHM / CM

Restivity is a test of the fluid’s resistance to conduct electrical potentials and thus

protect the valves from electrical erosion.

12.2.5 Fluid Conditioning

It cannot be emphasised too strongly that the key to a long fluid life is to keep the

fluid clean and dry and to maintain a low level of acidity. This is normally

achieved in situ by using solids treatment and / or vacuum dehydration to remove

acid degradation products and water. The equipment should preferably be a fixed

part of the system, but portable units can be used to conditin the fluid on a batch

basis. The solids treatment system should be of the correct size for the volume of

fluid if the system a 1.5-3% w/w ratio of absorbant solids: fluid is usually

adequate where 5-10% of the fluid volume is circulated every hour. If the

treatment system is too small, it will obviously be diffcult to maintain the

necessary low levels of acidity. All solids treatment systems should contain a 0.5

pm (nominal) fine filter immediately downstream of the solids filter in order to

remove fine particles which might otherwise spread throughout the system.

Although fullers earth has been used successfully for many years to treat

natural or coal tar phosphates, some depositin problems have arisen when this type

of sold has been used in conjunction



PAGE 140

with synthetic phosphates based on alkylated phenols. The deposists have

invariably been associated with high fluid acidities and high carbonate levels in

the fullers earth. The high acid levels can result in the production of insoluble

polymeric phosphates and leads to the formation of soluble and insoluble calcium

and magenesium salts through interaction with the carbonate content of the fuller's

earth. Since the fullers' earth is a naturally occring product, it has a cariable

composition, and on occasion high carbonate levels cannot be avoided. The

problem can be virtually eliminated by using activated alumina, a synthetic

product of constant composition.

In the presence of acid, hydrolysis of the fluid is autocatalytic, so it is

important to keep the acidity level as low as possible. It is recommended that the

acidity is not allowed to increase by more than 0.1 mg KOH / g above the limit

on the fresh fluid specified by the turbine manufacturer. Usually this means a

maximum service acidity of 0.2 mg KOH / g. The absorbant solid cartridges

should be replaced when the limit is exceeded and the acidity is showing a steady

increase. Alternatively, the solid can be replaced at regular intervals (for example

every 6 months) assuming that the acidity is monitored between filter changes, so

the reason for abnormal values should be investigated as soon as possible.

Absorbant solid filters can also become clogged by small particles and the

filter cartridges should be changed if the pressure drop across the filter exceeds the

manufactures recommendations (normally about 2 bar).

Since the fullers earth or activated alumina will rapidly absorb moisture

when exposed to the atmosphere, the replacement elements should preferably be

supplied and stored in air-tight plastic bags. Immediately before use, the solids

should be dried at 110oc for at least 12 hours. As moisture is quickly reabsorbed, if

the dried solids are left at ambient temperature for morre one hour after drying, the

filter element should be cooled to 20-30oc in the drying oven and immediately

transferred to the filter. A wet solid will not only be less effective in removing

degradation products, but it may also release water into the fluid and thus promote

hydrolysis.



When the turbine is shut down and the fluid allowed to cool, condensation

may occur in the tank, with the resulting hydrolysis of the tester. For outage

periods of longer than 1 week it is recommended that if possible the fluid be

conditioned each week for a sufficient time to ensure at least one passof the fluid

charge through the absorbant solid and the tank exhauster be operated during this

period.

Samples of fluid should be checked for acidity and water content at

intervals of (say) two weeks in order to ensure that the recommended limits are not

exceeded. If it is not possible to circulate the fluid through tje absorbant solid,

samples should still be taken at similar intervals for analysis, and if the acidity

rises to 0.3-0.5 mg KOH / g, conditioning should be started as soon as possible in

order to reduce the acidity to below the recommended limits before turbine

operatin resumes. If the acidity exceed 0.6 mg KOH / g it may be difficult or even

impossible to recondition the fluid charge and it may have to be replaced.

In view of the possible formation of metal salts by chemical reaction

between acidic degradation products and the fullers' earth or, much less likely,

alumina, their presence should be monitored at regular intervals.

Some solids can remove certain additives, e.g. the antifoam, from the

fluids, in the event of unsatisfactory fluid performance users are advised to check

with the fluid sppliers as to whether the additive levels should be restored.



PAGE 141

COMPARISON OF 500 MW AND 200 / 210 MW TURBINES (A) CONSTRUCTION 500 MW (KWU) 200 / 210 MW (KWU) 200 / 210 (LMW)

1. 3 cylinder reheat condensing turbine

3 cylinder reheat condensing turbine.

3 cylinder reheat condensing turbine.

2. Single flow HP turbine with 18 reaction stages

Single flow HP turbine with 25 reaction stages

Single flow HP turbine with 12 impulse stages

3. Double flow IP turbine with 14 reaction stages per flow

Double flow IP turbine with 20 reaction stages each flow

Single flow IP turbine with 11 impulse stages

4. Double flow LP turbine with 6 reaction stages each flow

Double flow LP turbine with 8 reaction stages each flow

Double flow LP turbine with 4 impulse Reaction stages (45% Reaction in last stage) each flow.

5. 4 Main stop and control valves (combined)

2 Main stop and control valves (combined)

2 Main stop valves(ESVs) & 4 control valves

6. 4 Reheat stop and control valves (combined)

2 Reheat stop and control (combined)

2 Interceptor valves and 4 control valves (CVs).

7. 2 Swing check NRVs in cold reheat line

2 Swing check NRVs in cold reheat line

2 Lift check NRVs in cold reheat line.

8. 2 Bypass stop and control valves 2 Bypass stop and control valves

2 Bypass stop and control valves

500 MW (KWU) 200 MW (KWU) 200 MW (LMW) 9. Rated steam flow 1568 T / hr. 670 T/hr. 670 T/hr. 10. Circulating water 54,000 m3 / hr 28,570 M3/hr 27,000 M3/hr 11. Type of Governing Hydraulic Hydraulic Hydraulic Electro-hydraulic Electro-hydraulic Electro-hydraulic (throttle) (throttle) (throttle) 12. No. of bearings 4 4 5 (Turbine only) 13. Rated speed 3000 r.p.m. 3000 r.p.m. 3000 r.p.m.

14. Max. Speed notime 3090 r.p.m. 3090 r.p.m. 3030 r.p.m.

limitation (51.5 Hz) (51.5 Hz) (50.5 Hz) 15. Min. speed no time 2850 r.p.m. 2850 r.p.m. 2940 r.p.m. limitation (47.5 Hz) (47.5 Hz) (49 Hz)

(B) * BHEL is considering to lower it to 2910 r.p.m. (48.5 Hz)

16. Permissible below 47.5 Hz and 51.5 Hz in life time (Total)

2 Hrs. 2 Hrs. BHEL is considering to allow 5 minutes continuous operation between 48 Hz and 48.5 Hz not exceeding 3 hrs. in total life.



PAGE 142 17. Critical Speed r.p.m. 900, 1548, 1896

2700, 3858 1200, 1600, 1900 1585, 1881, 2015

18. Speed exclusion range at operation without load

400-2850 r.p.m. 700-2850 r.p.m. 1100-2600 r.p.m.

19. Overspeed trip 3350 r.p.m. 3330 r.p.m. 3330 r.p.m. 20. Barring Gear Speed (r.p.m.) 240 / 210 (r.p.m.)

cut-in / cut-out (Hydralic)

540 / 500 cut-in / cut-out (Hydralic)

3.4 r.p.m. (motorised)

(C) STEAM PARAMETERS 21. Initial steam Pressure 170 Kg / cm2 147 Kg / cm2 130 Kg / cm2 (abs) (abs) (abs) 22. Initial steam temperature 537oc 535oc 535oc 23. Before HP 1st stage pressure

151.8 Kg / cm2 132.6 Kg / cm2 91 Kg / cm2

(abs) (abs) after CURTIS WHEEL (I.e. after regulating stage)

24. HP cylinder exhaust pressure

45 Kg / cm2 39.2 Kg / cm2 27 Kg / cm2

(abs) (abs) (abs) 25. HP cylinder exhaust temp. 342.5oc 343oc 327oc 26. IP cylinder stop valve 40.5 Kg / cm2 34 Kg / cm2 35 Kg / cm2 intel pressure. (abs) (abs) (abs) 27. IP cylinder stop valve 537oc 535oc 535oc intel temp. 28. Bleed Steam point (stages of turbine)

1 Nos. -CRH 2 Nos. - IPT 1 Nos. Cross around pipe between IPT & LPT, 2 Nos. LPT

25,36,45,48, 50,52

9,12,15,18,21,23,25

29. Extraction valves 500 MW (KWU) 200 MW (KWU) 200 MW (LMW) a. Extraction 1- No Valve 1 swing No Valve 1 swing No Valve 1 NRV b. Extraction 2- check valve with

Aux. Actuator check valve with Aux. Actuator

(lift check type) with Aux. actuator

c. Extraction 3- 1 swing check valve with Aux.actuator

1 swing check valve with Aux. Actuator

1 NRV (lift check type) with Aux. Actuator

d. Extraction 4,1- 2 swing check valves with Aux. Actuator

Extraction-4 1 swing check valve with Aux. Actuator


4,2 e. Extraction 5- 1 swing check valve

with Aux.actuator

1 swing check valve with Aux. Actuator


f. Extraction 6- No valve No valve 2 NRVs (lift check type) with Aux. Actuator

g. Extraction 7- 1 NRV (lift check type) with Aux. Actuator



PAGE 143 30. a.Extraction 7 - 39 Kg / cm2 Pressure / temp. (abc) 377oc b.Extraction 6 - 45 Kg / cm2 (abs) 39 Kg / cm2 (abs) 26 Kg / cm2 (abs) Pressure / temp. 537oc 537oc 537oc c.Extraction 5 - 19.5 Kg / cm2 (abs) 16.7 Kg / cm2 (abs) 12 Kg / cm2 (abs) Pressure / temp. 428oc 433oc 433oc d.Extraction 4 - 7.57 Kg / cm2 (abs) 7.06 Kg / cm2 (abs) 6.5 Kg / cm2 (abs) Pressure / temp. 302oc 316oc 368oc e.Extraction 3 - 2.76 Kg / cm2 (abs) 2.37 Kg / cm2 (abs) 2.8 Kg / cm2 (abs) Pressure / temp. 197.8oc 200oc 252oc f.Extraction 2- 1.42 Kg / cm2 (abs) 0.86 Kg / cm2 (abs) 1.3 Kg / cm2 (abs) Pressure / temp. 197oc 138.8oc 172oc g.Extraction 1- 0.286 Kg / cm2 (abs) 2.216 Kg / cm2 (abs) 0.3 Kg / cm2 (abs) 31. LP cylinder exhaust 0.884 Kg / cm2 (abs) 0.118 Kg / cm2 (abs) 0.09 Kg / cm2 (abs) Pressure / temp. 43oc 49oc 44oc 32. Low vacuum trip 0.31 Kg / cm2 (abs) 0.3 Kg / cm2 (abs) 0.3 Kg / cm2 (abs) (Hyd. & Electical) (540 mm vac) 33. Bypass trip 0.61 Kg / cm2 (abs) 0.6 Kg / cm2 (abs) 0.41 Kg / cm2 (abs) (450 mm vac.)



Power Plant Familiar is at Ion v- III

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