Power Plant Familiar is at Ion v- III

Page-1

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 Page-2 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 Page-3 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. page-4

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. PAGE 5 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 PAGE 6 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.


PAGE 7 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.

PAGE 8 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. condition for a unit on a T.S. diagram are indicated in PAGE 9


Ideal

PAGE 10

PAGE 11 (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


PAGE 12 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 – C02) = h0 – h1 PAGE 13 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 - V12) = 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 _ C22) + (V22 _ V12) 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. PAGE 14 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.

PAGE15


PAGE18 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. PAGE 19

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

PAGE 20 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-molybdenumvanadium steel, or more recentn 0 3 112 per cent nickel-chromium-molybdenumvanadium 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 bypass 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 onload 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 variablespeed 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

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 already in water

- 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 Ushaped 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 interlocked 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

2.

a)

Turbine oil 14

b)

Mobil DTE medium

a)

Specific gravity at 50'c

b)

Kinematic viscosity at 50'c 28 CS

c)

Neutralisation number

d)

Flash point 201'c (min)

e)

Pour point

f)

Ash percentage by weight 0.01%

g)

Mechanical impurities

0.852

0.2

-6.60c (max)

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 electrohydraulic transducer which closes the control valves when generator circuitbreaker 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-upPistons" 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 electrohydraulic 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 nonreturn 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


Over 100 microns


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)

1.

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

200 / 210 (LMW) 3 cylinder reheat condensing turbine. 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 500 MW (KWU) 200 MW (KWU) 1568 T / hr. 670 T/hr. 54,000 m3 / hr 28,570 M3/hr Hydraulic Hydraulic Electro-hydraulic Electro-hydraulic (throttle) (throttle) 4 4

2 Bypass stop and control valves 200 MW (LMW) 670 T/hr. 27,000 M3/hr Hydraulic Electro-hydraulic (throttle) 5

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 15. Min. speed no time limitation

(51.5 Hz) 2850 r.p.m. (47.5 Hz)

(51.5 Hz) 2850 r.p.m. (47.5 Hz)

(50.5 Hz) 2940 r.p.m. (49 Hz)

9. Rated steam flow 10. Circulating water 11. Type of Governing

12. No. of bearings (Turbine only) 13. Rated speed

(B) 16. Permissible below 47.5 Hz and 51.5 Hz in life time (Total)

* BHEL is considering to lower it to 2910 r.p.m. (48.5 Hz) 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 400-2850 r.p.m.

18. Speed exclusion range at operation without load 19. Overspeed trip 3350 r.p.m. 20. Barring Gear Speed (r.p.m.) 240 / 210 (r.p.m.) cut-in / cut-out (Hydralic)

1200, 1600, 1900 1585, 1881, 2015 700-2850 r.p.m.

1100-2600 r.p.m.

3330 r.p.m. 540 / 500 cut-in / cut-out (Hydralic)

3330 r.p.m. 3.4 r.p.m. (motorised)

(C) STEAM PARAMETERS 21. Initial steam Pressure

170 Kg / cm 2

147 Kg / cm2

130 Kg / cm 2

22. Initial steam temperature

(abs) o 537 c

(abs) o 535 c

(abs) o 535 c

23. Before HP 1st stage pressure

151.8 Kg / cm

2

(abs)

24. HP cylinder exhaust pressure

45 Kg / cm

2

25. HP cylinder exhaust temp. 26. IP cylinder stop valve

40.5 Kg / cm

intel pressure. 27. IP cylinder stop valve

(abs) o 537 c

29. Extraction valves a. Extraction 1b. Extraction 2-

c. Extraction 3-

d. Extraction 4,1-

4,2 e. Extraction 5-

f. Extraction 6-

2

(abs)

(abs) o 342.5 c

intel temp. 28. Bleed Steam point (stages of turbine)

132.6 Kg / cm

2

(abs) o 343 c

1 Nos. -CRH 2 Nos. - IPT 1 Nos. Cross around pipe between IPT & LPT, 2 Nos. LPT 500 MW (KWU) No Valve 1 swing check valve with Aux. Actuator

34 Kg / cm

2

after CURTIS WHEEL (I.e. after regulating stage)

39.2 Kg / cm

2

91 Kg / cm

27 Kg / cm

2

(abs) o 327 c 2

35 Kg / cm

2

(abs) o 535 c

(abs) o 535 c

25,36,45,48, 50,52

9,12,15,18,21,23,25

200 MW (KWU) No Valve 1 swing check valve with Aux. Actuator

200 MW (LMW) No Valve 1 NRV (lift check type) with Aux. actuator

1 swing check valve with Aux.actuator

1 swing check valve with Aux. Actuator 2 swing check valves Extraction-4 with 1 swing check Aux. Actuator valve with Aux. Actuator

1 NRV (lift check type) with Aux. Actuator

1 swing check valve with Aux.actuator

1 swing check valve with Aux. Actuator


No valve

No valve

2 NRVs (lift check type) with Aux. Actuator 1 NRV (lift check type) with Aux. Actuator

g. Extraction 7-



PAGE 143 30. a.Extraction 7 -

39 Kg / cm o

Pressure / temp. b.Extraction 6 Pressure / temp. c.Extraction 5 Pressure / temp. d.Extraction 4 Pressure / temp. e.Extraction 3 Pressure / temp. f.Extraction 2Pressure / temp. g.Extraction 131. LP cylinder exhaust Pressure / temp.

2

(abc) 377 c 45 Kg / cm 2 (abs) o

39 Kg / cm 2 (abs) o

537 c

o

537 c 2

19.5 Kg / cm (abs) o

537 c 2

2

2

16.7 Kg / cm (abs)

12 Kg / cm (abs)

433o

433 c

o

c

428 c

26 Kg / cm2 (abs)

2

2

7.57 Kg / cm (abs)

7.06 Kg / cm (abs)

6.5 Kg / cm (abs)

302oc

316oc

368oc

2


2


197.8 c

o

200 c 2

2

2.8 Kg / cm (abs) 252 c

2

2

1.42 Kg / cm (abs)

0.86 Kg / cm (abs)

1.3 Kg / cm (abs)

197oc

138.8oc

172oc

2

2.216 Kg / cm (abs)

2

0.118 Kg / cm (abs)

0.286 Kg / cm (abs) 0.884 Kg / cm (abs) o

2 2

o

43 c

44 c

2

0.3 Kg / cm (abs)

2

0.6 Kg / cm (abs)

0.31 Kg / cm (abs)

(Hyd. & Electical) 33. Bypass trip

0.61 Kg / cm (abs)

2


49 c

32. Low vacuum trip

2

0.3 Kg / cm (abs)

2

2

2

0.3 Kg / cm (abs) (540 mm vac) 2 0.41 Kg / cm (abs) (450 mm vac.)


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