combined gas & steam turbine power plants

Chapter 12 ....................................................... Conclusions 353

............................................................ Conversions.. 355

Symbols Used.. .......................................................... 357

Indices Used ............................................................. 359

Appendix 1 ................................................................ 3f3

................................. Definition of Terms and Symbols 371

Bibliography ............................................................. 37'7

Chapter I

INTRODUCTION

rature has often suggested combining two or more ther- ithin a single power plant. In all cases, the inten- crease efficiency over that of single cycles. Thermal n be combined in this way whether they operate e or with differing working media. However, a com-

cycles with different working media is more inter- se their advantages can complement one another.

the cycles can be classed as a "topping" and a "bot- . The first cycle, to which most of the heat is sup- the "topping cycle." The waste heat it produces in a second process which operates at a lower

vel and is therefore referred to as a "bottoming

on of the working media makes it possible to process that makes optimum thermodynamic the upper range of temperatures and returns

nvironment at as low a temperature level the "topping" and "bottoming" cycles are

h e , only one combined cycle has found e combination gas turbinelsteam turbine lants of this type have burned generally y-liquid fuels or gases.)

2 COMBINED CYCLE GAS & STEAM TURBINE POWER PLANTS INTRODUCTION 3

Fig. 1 is a simplified flow diagram for an installation of this It therefore is quite reasonable to use the steam process for type, in which an open-cycle gas turbine is followed by a steam process. The heat given off by the gas turbine is used to gen- ine power plants were not more widely used even earlier erate steam. clearly been due to the historical development of the gas

ine. Only in recent years have gas turbines attained inlet Other combinations are also possible, e.g., a mercury vapor

process or replacing the water with organic fluids or ammonia.

The mercury vapor process is no longer of interest today since even conventional steam power plants achieve higher efficienc-

orld totals more than 30,000 MW.

ies. Organic fluids or ammonia have certain advantages over water in the low temperature range, such as reduced volume flows, no wetness. However, the disadvantages, i.e., development costs, environmental impact, etc., appear great enough to prevent their ever replacing the steam process in a combined-cycle power plant. The discussion that follows deals mainly with the combination of an open-cycle gas turbine with a waterhteam cycle. Cer- tain special applications using closed-cycle gas turbines will also be dealt with briefly.

Why has the combination gas turbinehteam turbine power plant, unlike other combined-cycle power plants, managed to find wide acceptance? Two main reasons can be given:

@ It is made up of components that have already proven themselves in power plants with a single cycle. Devel- opment costs are therefore low.

The steam process uses water, which is likewise inexpensive flow diagram of a combination gas turbinelsteam turbine power

and widely available, but better suited for the medium and low temperature ranges. The waste heat from a modern gas turbine

4. Steam turbine has a temperature level advantageous for a good steam process. 5. Condenser

6. Fuel supply

Chapter 2

MODYNAMIC PRINCIPLES OF PLANT

Considerations

t efficiency is the maximum efficiency of an ideal

Carnot efficiency

Temperature of the energy supplied

Temperature of the environment

the efficiencies of real processes are lower since s involved. A distinction is drawn between energetic losses. Energetic losses are mainly heat n and convection), and are thus energy that is ess. Exergetic losses, on the other hand, are in- sed by irreverisible processes in accordance with of thermodynamics [I].

o major reasons why the efficiencies of real pro- er than the Carnot efficiency:

erature differential in the heat being supplied ry great. In a conventional steam power plant,

e maximum steam temperature is only about

6 COMBINED CYCLE GAS & STEAM TURBlNE POWER PLANTS

810K (980°F), while the combustion temperature in the boiler is approx. 2000 K. Then, too, the temperature of the waste heat from the process is higher than the ambient temperature. Both heat exchange processes cause losses.

The best way to improve the process efficiency is to reduce these losses, which can be accomplished by raising the maximum temperature in the cycle, or by releasing the waste heat at as low a temperature as possible.

The interest in combined-cycles arises particularly from these two considerations. By its nature, no single cycle can make both improvements to an equal extent. It thus seems reasonable to combine two cycles: one with high process temperatures, and the other with a good cold end.

In an open-cycle gas turbine, the process temperatures attainable are very high because its energy is supplied directly to the cycle without heat exchangers. The exhaust heat temperature, however, is also quite high. In the steam cycle, the maximum process temperature is not very high, but the exhaust heat is returned to the environment on the cold end at a very low temperature.

Combining a gas turbine and a steam turbine thus offers the best possible basis for a high-efficiency thermal process (Table 2-1).

The last line in the table shows the "Carnot efficiencies" of the various processes, i.e., the efficiencies that would be attainable if the processes took place without internal exergetic losses. Although that naturally is not the case, this figure can be used as an indicator of the quality of a thermal process. The value shown makes clear just how interesting the combined-cycle power plant is when compared to the single-cycle processes. Even a sophisticated installation such as a reheat steam turbine power plant has a theroretical Carnot efficiency 10 to 15 points lower

MODYNAMIC PRINCIPLES OF THE COMBINED-CYCLE PLANT 7

hat of a combined-cycle plant. On the other hand, the ex- c losses in the combined cycle are higher because the tem-

e differential for exchanging heat between the exhausts e gas turbine and the waterlsteam cycle is relatively great. s clear why the differences between the actual effici- ttained by a combined-cycle power plant and the other s are not quite that large.

n by Fig. 2-1, which compares the temperaturelentropy f the four processes, the combined cycle best utilizes ature differential in the heat supplied, even though additional exergetic loss between the gas and the

-1: Thermodynamic Comparison of Gas Turbine, Steam Turbine, and Combined-Cycle Power Plants

Reheat Reheat Power Plant

ed, in K 950 - 1000 640 - 700 550 - 630 950 - 1000 (1250 - 1340) (690 - 800) (530 - 675) (1250 - 1340)

500 - 550 320 - 350 320 - 350 320 - 350 (440 - 530) (115 - 170) (115- 170) (115- 170)

10 COMBINED CYCLE GAS & STEAM TURBINE POWER PLANTS

Combining these two equations yields:

2.2.1 The Effect of Additional Firing in the Waste Heat Boiler on Overall Efficiency

Substituting Equations (4) and (7) into Equation (2), one obtains:

VGT GT + VST (QSF + Q G T [ I - VGTI) TIK =

(8) QGT + QSF

Additional firing in the waste heat boiler improves the overall efficiency of the combined-cycle installation whenever:

Differentiation of Equation (8) produces the inequality:

a V S T @GT + QSF) - VST QSF + [= QGT (1 - VGT)] - (10)

G ~ + b F ) - V ~ Q G T (1- vGT)J > O

This yields:

RMOD YNAMlC PRlNClPLES OF THE COMBINED-CYCLE PLANT I I

nee the second term of the inequality is equal to I(, the in- ality reduces to:

(12)

)I is none other than the heat to the steam cycle. The formula thus becomes:

(13)

tion (13) means that increasing the additional firing im- the efficiency of the combined-cycle plant only if it im- the efficiency of the steam process. The greater the

nee is between the efficiencies of the combined-cycle and er the temperature is of the heat

the steam process, the more effective that improve- Ll be. For that reason, additional firing is becoming less

cy of the combined-cycle instal- eases far more rapidly than that of the steam process, y increasing the difference (TK - ST). In view of the

is generally better to burn the modern gas turbine, because the heat is supplied to

ess at a temperature level higher than that in the steam

bined-cycle installations with firing are discussed in more detail in Section 3.2 below.

12 COMBINED CYCLE GAS & STEAM TURBINE POWER PLANTS MODYNAMIC PRINCIPLES OF THE COMBINED-CYCLE PLANT 13

2.2.2 Efficiency of Combined-Cycle PlanLs without 2-2: Allowable Reduction in Steam Process Efficiency Additional Firing in the Waste Heat Boiler as a Function of Gas Turbine Efficiency (Steam

Without additional firing, Equation (8) can be written as fol- process efficiency = 0.25)

lows: (14)

~ G T 0.2 0.3 0.4 VGT - QGT + VST . QGT (1 - TIGT) = sGT + (1 - sG VK =

QGT - - *ST 0.94 1.07 1.25

Differentiation makes it possible to estimate the effect that a change in efficiency of the gas turbine has on overall efficiency:

a 7 7 ~ = 1 + --- a VST (1 - VGT) - VST (15) in the steam cycle. But a gas turbine with a maxi-

a VGT a VGT ciency still does not provide an optimum combined- nt. For example- with a constant turbine inlet

Increasing the gas turbine efficiency improves the overall ef- ure- a gas turbine with a very high pressure ratio at-

ficiency only if: her efficiency that a machine with a moderate pres- wever, the efficiency of the combined-cycle plant

a r]K (16) nd machine is sigmficantly better because the steam > 0 a VGT ollows operates far more efficiently with the higher temperature and produces a greater output.

From Equation (15) one obtains: I ows the efficiency of the gas turbine alone as a

afl < 1 - ssr (17) e turbine inlet and exhaust temperatures. The max-

-- 1 - ~ ] S T cy is reached when the exhaust gas temperatures

~ G T (A low exhaust temperature means a high pres- Improving the gas turbine efficiency is helpful only if it does

not cause too great a drop in the efficiency of the steam Process.

ST ws the overall efficiency of the combined-cycle

Table 2-2 shows the maximum allowable reduction- - y. Compared to Fig. 2-2a, the optimum point has as a function of the gas turbine efficiency. ~ G T higher exhaust temperatures from the gas tur-

omical considerations, present-day gas turbines This table indicates that the higher the efficiency of the gas imized with respect not to efficiency but to max-

turbine, the greater may be the reduction in efficiency of the nsity. Fortunately, this optimum coincides fairly steam process. The proportion of the overall output being pro- h the optimum efficiency of the combined-cycle vided by the gas turbine increases, reducing the effect of lower It- most of today's gas turbines are optimally

ined-cycle installations.


Gas turbines of a more complicated design, i.e., with inter- mediate cooling in the compressor or recuperator, are less suitable for combined cycles. They normally have low exhaust gas temperatures, so that the efficiency of the steam turbine can only be low. We shall not discuss a reheat gas turbine here since this type of machine has disappeared from the market due to its complexity. a

In summary, it may be said that:

The gas turbine with the highest efficiency does not necessarily produce the best overall efficiency of the combined-cycle plant. The turbine inlet temperature is a far more important factor.

Similar considerations also apply with regard to the efficiency of the steam cycle. These, however,are less important because the gas turbine is generally the "standard machine." The exhaust heat available for the steam process is thus a given, and the problem lies only in its maximum conversion into mechanical energy (refer on this point to Section 2.3.)

b

Efficiency of Gas Turbines in cornbined-Cycle the Turbine Inlet and Exhaust Gas T~~~~~~~~~~~

18 COMBINED CYCLE G A S & STEAM TURBINE N I W E R H A N T S SYSTEM LAYOUTS 19

In addition to this physical limitation, there is also a chemical- Combined-cycle plants without additional firing often are made limitation on energetic use of the exhaust gases imposed by low P of several gas turbines and waste heat boilers that supply temperature corrosion. This corrosion, caused by sulphur, oc- to a single steam turbine. In the following, we generally curs whenever the exhaust gases are cooled below a certain tern- only of one gas turbine and one waste heat boiler, but perature, the sulphuric acid dewpoint. youts can also be adapted for several gas turbines. Because

implest system is typical of all, it has been discussed more In a waste heat boiler, the heat transfer On the flue gas side tail, and the other possibilities have then been derived from

is not as good as on the steam or water side. For that reason1 the surface temperature of the pipes on the flue gas side is aP- proximately the same as the water or steam temperature. If these ingle-Pressure System pipes are to be protected against an attack of low temperature corrosion, the feedwater temperature must remiin approximately plest arrangement for a combined-cycle plant is a single- as high as the acid dewpoint. Thus, a high stack temperature for the flue gases does no good if the temperature of the feed- f one or more gas turbines with a single-pressure waste water is too low (refer also to Section 5.2). Low temperature corrosion can occur even when burning fuels containing no sul- and a single-stage feedwater preheater in the de- phur if the temperature drops below the water dewpoint. e Steam for the deaerator is tapped from the steam

3.1 Combined-Cycle Plants without Additional Firing ste heat boiler consists of three paas: In combined-cycle plants without additional firing, all the fuel

is burned in the gas turbine. The steam turbine then utilizes the dwater preheater (economizer), which is by the flue gases;

exhaust heat from the gas turbine, with no additional source of thermal energy. This type of combined-Cycle plant is already in widespread use because it is simple and inexpensive and high efficiencies can be attained with modern gas turbines.

ural circulation.

combined-cycle plants is quite large because attempts have be made to improve the quality of the heat exchange between t Ble-Pressure System

flue gas and the water or steam by using complex systems. s the heat balance in a typical single-pressUre has led to systems that utilize the exhaust heat well both e getically and energetically. enerator aPProx. 35 kgls (277,200 lb/hr) steam

with an output of 35 MW. Because of the good


Figure 3-1 SYSTEM LQYOUTS 21

- HEAT TRANSFER

Fig. 3-1: TemperatureEXeat Diagram: Ideal Heat Exchange

22 COMBINED CYCLE G A S & STEAM TURBINE POWER PLANTS

Figure 3-3

SYSTEM LAYOUTS 23

Fig. 3-3: Flow diagram of the single-pressure system

1 Compressor 6 Economizer ll Feedwater tank/ 2 Gas turbine 7 Boiler drum deaerator 3 Bypass stack 8 Steam turbine 12 Feedwater pump 4 Superheater 9 Condenser 13 Condensate pump 5 Evaporator 10 Steam bypass

24 C O M B I N E D CYCLE G A S & STEAM TURBINE POWER P U N T S SYSTEM U Y O U T S 25

hewn in Fig. 3-7, the smallest temperature differ- river-water cooling system, pressure in the condenser is 0-04 between the water and the exhaust gases in the bar (0.58 psia), resulting in a gross efficiency of the irlstauation omizer is on the warmer end of the heat ex-

of 45% (Table 3-1, page 30). ger. 'I'hat means: the amount of steam production lble does not depend on the feedwater tempera- . In a conventional steam generator, on the other

Noteworthy is the poor energetic utilization of the exhaust the sr.nallest temperature difference is on the heat from the gas turbine. Together with the relatively low live end of the economizer because the water flow steam data, this produces a fairly modest efficiency in the stearn larger in proportion to the flue gas flow. AS a process. Fig. 3-5 shows the energy flow. t, the WnoUnt of steam production possible de-

s on the feedwater temperature.

45% of the thermal energy supplied is converted into electri- wS two examples of conventional steam generators cal energy. The rest is removed in the condenser (28.3%) or g feed-water temperatures. ~t is obvious that with

through the stack (25.2%) or is lost elsewhere (1.5%). rmce in temperature at the end of the econo-

Fig. 3-6 shows the exergy flow of the Same plant. The heat t available for evaporation and superheating is

that has to be removed in the condenser is only about half that greater where the feedwater temperature is high-

0s a conventional steam power plant of the same size. m ~ o u n t of live steam produced by a conventional increased by raising the feedwater temperature.

One significant difference between a Conventional steam plant and the steam process in a combined-cycle plant lies in the boil rnbient Conditions

f eedwater preheating. A conventional Steam plant attains a ter efficiency if the temperature of the feed-water is brou s here only the effect that different ambient to a high level by means of multi-stage preheating. In a c on the design point for the installation. HOW

bined-cycle power plant, however, the boiler feedwater m sioned combined-cycle plant behaves will be be as cold as possible, with the limit determined by low te on 7, Operating and Part-Load Behavior. Those perature corrosion: the temperature of the water must not e valid, however, only for the steam turbine significantly below the dewpoint for sulphuric acid. There a two reasons for this difference:

mbient conditions. This can be justified 0 Normally, a conventional steam generator is equippe

with a regenerative air preheater that can further ut gas turbine that has been optimized for

ilize the energy remaining in the flue gases after the f 15°C (59°F) does not look significantly economizer. There is nothing like that in a waste he t has been designed for, say, 40°C (104°F). boiler, so that the energy remaining in the ~xhaus t ng a new machine would thus not be gases after the economizer is lost.

26 COMBINED CYCLE GAS & STEAM TURBINE POWER PLANTS SYSTEM LAYOUTS 27

Figure 3-5

30,l %

25,2 1 4 , 9 7 4

Fig. 3-5: Energy Flow Diagram for the Single-Pressure Combined-Cycle Plant

Q Energy input Diagram of the Single-Pressure Combined-Cycle Plant V1 Loss in condenser V2 Loss in stack V3 Loss due to radiation in waste heat boiler V4 Loss in flue gas bypass V5 Loss in generator and radiation, gas turbine V6 Loss in generator and radiation, steam turbine GT Electricity produced in the gas turbine ST Electricity produced in the steam turbine

e steam turbine ste heat boiler

SYSTEM LQYOUTS 29


Table 3-1: Main Technical Data of the Single-Pressure Increasing the air temperature reduces the density of Combined-Cycle Plant the air, and thereby reduces the air mass flow drawn

y the compressor increases in Gas turbine output take temperature (in K), without Steam turbine output being a corresponding increase in the output

Station service power required

Net power output of plant 101 500 kW capacity of the turbine re- 228 000 kW pressure before the turbine is re-

Thermal energy supplied (Diesel fuel) decreases as the air Efficiency of gas turbine in reduces the pressure ra-

157 000 kW same principle applies in- Heat contained in exhaust gases compressor, but because its Utilization rate for waste heat energy* an of the turbine, the total bal- Efficiency of the steam process

Gross efficiency of the plant ows this change in a temperaturelentropy diagram. s that the exhaust gas temperature becomes higher

ses. This is because the turbine pres- * 100% utilization if the exhaust gases are cooled down to 15 OC (59 OF duced while the inlet temperature remains con-

avior of the exhaust gas temperature explains The situation is different on the steam end of the steam tu that the air temperature has on the efficiency

cycle plant differs from that which it has on the sure of, say, 0.2 bar (2.9 psia) can no longer function Proper he gas turbine alone. if the pressure is only 0.04 bar (0.58 psia).

e efficiencies of the gas turbine and plant as a function of the air temperature,

the air temperature, air pressure, and cooling water temper aining otherwise unchanged. As it ture. The relative humidity is important only if the water the air temperature even has a slightly cooling the condenser is recooled in a wet Cooling tower. ficiency of the combined-cycle plant,

erature in the gas turbine exhaust rai- Air Temperature team process (Fig. 3-1 1) enough to more

There are three reasons why the air temperature has a 1 uced efficiency of the gas turbine influence on the power output and efficiency of an open-c

urprising when one remembers the Car- n (1)). The rise in the final temperature

SYSTEM U Y O U T S 33

34 COMBINED CYCLE GAS & STEAM TURBINE POWER PUNTS

Figure 3-11

-20 -10 0 + I 0 + 2 0 3 0 40 5 0

A I R TEMPERATURE

Fig. 3-11: Relative Efficiency of the Steam Process in Combined-Cycle Plants Function of the Air Temperature Cooling water temperature 20°C (68°F)

SYSTEhd LAYOUTS 35

ression causes a slight increase in the average temper-

es only if the temperature of the water cooling the

air-cooled condenser, the efficiency of the steam nges because the condenser pressure is now different.

nges with the air temperature when the cooling

e case with direct air-cooled condensation.

m its efficiency. Here the reduced flows of air es play a more important role than the exhaust

how the power outputs of the gas turbine and le plant change depending on the air temper-

ite Elevation

e air pressure on the efficiency of a gas tur- ro if the temperatures remain unchanged. On

36 COMBINED CYCLE GAS & STEAM TURBINE POWER PLANTS SYSTEM LA YOUTS

taken in, which varies in proportion to the intake pressure a thereby also affects the flow of exhaust gas. The exhaust heat

well to the real situation, this then causes a similar variati in the power output from the steam turbine.

Because the power outputs of the gas turbine and the st turbine vary in proportion to the air pressure, the total po output of the combined-cycle plant varies correspondingly. efficiency of the plant remains constant, however, since the thermal energy supplied and the air flow are varying in p portion to the air pressure.

Cooling media for the Condenser

To condense the steam, a cooling medium must be used to car A I R TEMPERATURE

which has a high specific thermal capacity and good heat tra perature on the Efficiency of Combined-Cycle Plants

fer properties. Where water is in short supply, cooling can done in air in a wet cooling tower; where no water is availab an air-cooled condenser or a dry cooling tower are necessa

The temperature of the cooling medium affects the efficien

producing a greater useful enthalpy drop in the steam turbi

pressure as a function of the design temperature for the cooli medium. There are three different cases:

1 O C

with


Figure 3-13

-20 -10 0 + I 0 + 2 0 3 0 40 5

A I R T E M P E R A T U R E

Fig. 3-13: Effect of the Air Temperature on the Efficiency of Combined-Cycle with Direct Air-Cooled Condensation

SYSTEM LAYOUTS 39

SYSTEM LAYOUTS 41

42 COMBINED CYCLE GAS & STEAM TURBINE POWER PUNTS SYSTEM UYOUTS 43

@ direct water cooling e Steam Pressure water cooling, with water recooled in a wet cooling steam pressure does not tower . 3-17 shows how the ef-

@ direct air cooling on the live steam pres- is striking that the best efficiency is attained even while

The greatest vacuums are attained with direct water cooling, steam pressure is quite low. the least with direct condensation with air. In the comparison, it must also be borne in mind that the water temperature is gen an increased efficiency erally lower than that of the air. steam cycle due to the greater enthalpy gradient

. The rate of waste heat energy utilization in the For the wet cooling tower, a relative air humidity of 60% h , however, drops off sharply. The overall effici-

been assumed. am process is the product of the rate of energy

The Effect of the Most Important Design Parameters the efficiency of the waterlsteam cycle. There

on Power Output and Efficiency at approx. 30 bar (435 psia).

When dimensioning a combined-cycle plant, the gas t of energy utilization in design is generally a given, since the gas turbine is a sta oiler: the temperaturelheat diagrams are for dized machine. of 15 and 60 bar (203

e steam pressure, there The free parameters for the design involve the steam pr energy available for evaporation and super-

and it is mainly these that are discussed below. One mu ure is correspondingly forget, however, that the output of the steam turbine is int of the evaporator is the same in both approx. 30 to 40% of the total power output. Optimizati ce area of the heat exchanger is therefore the steam process can therefore only influence that porti result, the stack temperature at 15 bar is

Another important point: The efficiency of the steam r than at 60 bar, which means that the eing better utilized.

is always proportional to the output of the steam turbine, si in a plant without additional firing, the thermal energy supp lso greatly affects the to the steam process is a given. ved in the condenser (Fig. 3-19). The

en pressures are lower since a greater Live Steam Data removed from the exhaust gases and

The selection of the live steam data for a combined-cycle p at a lower efficiency. with a single-pressure system is a compromise between o mum energetic and optimum exergetic utilization of the ke it advisable to raise heat from the gas turbine. The main determining factor is e the thermodynamic optimum. This live steam pressure selected.

44 COMBINED CYCLE G A S & STEAM TURBINE POWER PLAN SYSTEM LAYOUTS 45

@ a reduction in the exhaust steam flow, or, if the si of the steam turbine remains unchanged, smaller e i~nproved efficiency of the waterlsteam cycle haust losses. ompensates for the slight drop in the rate of

@ a smaller condenser y utilization. Moreover, for the steam turbine, @ a reduction of the cooling water requirement e steam temperature means less erosion in the

use of the reduced water content in the steam). Especially in the case of power plants with expensive air

condensers, this can mean considerably lower costs. re of the gas turbine exhaust gas provides the

Live steam flows greater than that in the example sho in temperature is necessary between the ex- the optimum toward higher live steam pressures, since live steam in order to limit the size of the

over, too high a live steam temperature can rtionate increase in plant costs since a great

combined-cycle plants with several gas turbines to sele team turbine. In most cases, however, the steam pressure that is above the optimum. The reduced rature sets the limit for the live steam tem- flow that results makes it possible to employ piping and

of live steam is reduced. The optimum live steam pres depends on the total amount of live steam: increasing t ood rate of waste heat energy utilization, improves efficiency in the high pressure section of e feedwater should be kept as low as pos- turbine. With a larger volume flow, longer blades are re in the first row, which reduces the edging losses.

Live Steam Temperature in Section 3.1.1, preheating has been In contrast to the live steam pressure, raising the live

temperature always brings with it a slight increase in effic Id improve the efficiency but it has not (Fig. 3-20). There are two reasons for this improvement because the solutions shown in Sections increased superheating: arly better. Dividing preheating into sev-

mprove the rate of energy utilization in this improved thermodynamics of the cycle, e system, which is the greatest disadvantage increased steam turbine efficiency due to reduc m. Even with minimum feedwater temper- wetness in the low pressure section. mperature remains at approx. 200 O C (392

50 COMBINED CYCLE GAS & STEAM TURBINE POWER PUNTS

Figure 3-21

SYSTEM UYOUTS 51

nser Pressure

nce on the effici- py drop in the steam

anges sharply (Fig. 3-22). An increase in the pressure decrease in power output. However, plant costs are

nt of the Waste Heat Boiler

inch point") of the er, which affects the amount of steam generated -7). By reducing the pinch point, the rate of en-

However, the surface of the heat exchanger in- entially, which quickly sets a limit for the utiliza-

he waste heat boiler should be such that the s as low as possible. This nd efficiency of the gas the turbine. In present-

oss is approximately 0.8% for each 1% recovered in the steam

rate of recovery is 35%.

of the gas turbine exhaust gas is im- cle. If the turbine inlet ine with a higher ex-

Fig. 3-21: ~ f f ~ ~ t of the Feedwater Temperature trw on the r overall efficiency produces process and Rate of Waste Heat Energy Utilization tical cornpressor and

tfw = Feedwater Temperature [Other terms as in Fig. 3-17]

SYSTEM M Y O U T S 57

a Low

I-)


10% of that required for the version using w , however, is low compared to the improvement in ator itself can be of the natural circulation or the for

mplicated and accordingly low in cost. Even if the very high levels of sulphur, the feedwater can be

In this second design, it is sometimes possi sufficiently high a temperature without any re- arate low pressure drum. The feedwater tank the ficiency worth mentioning. as a low-pressure drum, resulting in a simp feed pumps or drum level controls are required. ows the temperature/heat diagram for the waste cause of the two-phase flow, special care he exhaust gases are cooled by approximately an designing the piping and the introduction of the O C (90 OF') in the preheating loop in order to warm. mixture into the feed-water tank. te to 130 O C (266 O F ) .

Example of a Single-Pressure Combined-Cycle Technical Data of the Single-Pressure Plant with a Preheating Loop ined-Cycle Plant with a Preheating Loop

This is shown in Fig. 3-27, using the same gas turbin the example for the simple single-pressure system (Fig.

the steam turbine 36 800 kW Table 3-3 lists the main technical data of this system 1 200 kW

equipped with a low pressure evaporator. 104 000 kW

Compared to the simple single-pressure system, it att 228 000 kW

nifieantly higher steam turbine output, improvi 30.0 % ciency by 2.5%. This is because in this case no steam 157 200 kW

from the turbine. As a result, the entire live ste 72.5 %

pand to the condenser pressure. But the larger volu 23.4 %

exhaust steam produced is a certain disadvantage si 46.1 %

mensions of the steam turbine exhaust and the cond

The increase in the amount of heat to be remove xhaust gases are cooled down to 15 O C (59 OF)

condenser is more than proportional to the increas output. The energy utilization rate of the waste heat tal Conditions

by about 15% while the power output from t increases only by 8%, since the additional exhaust heat ct the combined-cycle plant with a pre- is at a low temperature level. The rate for convertin ately the same way as the simple single- chanical energy (exergy) is therefore modest. The in &on 3.1.2). We will therefore no treat

60 COMBINED CYCLE GAS & STEAM TURBINE POWER PUNTS SYSTEM LAYOUTS

separately the various parameters that depend on the e ment.

Fig. 3-29 shows the effects that the temperatures of and the cooling water have on the power output and ef of the plant as a whole. It is obvious that a rise in air ature causes a reduction in power output and a s ment in overall efficiency. On the other hand, a high tem for the cooling water affects both parameters negativ

Effect of the Most Important Design Parameters on Power Output and Efficiency

The effect of most parameters is similar to that for the single-pressure system (See Section 3.1.1).

Live Steam Data

The effects of live steam pressure and Live steam tem on the efficiency of the steam cycle are practically the for a simple single-pressure system. The optimum live ste sure is at approximately the same level. Slight shifts t higher pressure can result due to a larger exhaust steam flow.

However, installing a preheating loop in the waste imposes a limit on the minimum live steam pressu seen from Fig. 3-30, the flue gas temperature af omizer drops when the live steam pressure falls. minimum temperature of the water in the boiler by the sulphuric acid dewpoint, the amount of use the preheating loop is reduced correspondingly.

If a high feedwater temperature is required, t pressure selected must not be too low. Otherwise a the preheating would have to be done in a low pressure

66 COMBINED CYCLE GAS & STEAM TURBINE POWER PLAN SYSTEM LAYOUTS 67

A low pressure preheater has a negative effect on the ptimum. In many cases, the low pressure evaporator process efficiency because less heat is recovered from t no great expense, produce more steam than required haust gases. However, the reduced wetness and exit 10s t the feedwater and that excess steam could be con- the turbine to a large extent compensate for that negative 0 mechanical energy if it were admitted into the tur-

suitable point. To do this, the steam turbine must Condenser Pressure m admissions: one for high pressure, and another

The effect of the condenser pressure on the efficiency ure steam (two-pressure turbine). steam process is similar to that in the simple single-pressu tern, but the change in efficiency is somewhat more prono ows a system of this type, further equipped with

because the exhaust steam flow is about 10 to 15% gre ure pre-heaters. This not only provides better util- waste heat as mentioned above, but also makes

Pinch Point of the Waste Heat Boiler dynamic use of the low pressure steam. A larger the low pressure steam can flow into the turbine

The effect that the pinch point of the waste heat pressure preheater, while the feedwater is be- on the efficiency of the steam process is similar to tha n the first section using low quality steam. ple single-pressure boiler (cf. Section 3.1.1). However tion of the pinch point affects not only the surface pressure steam reaches the turbine, it can be evaporator and the economizer but also that of the p ated. The thermodynamic advantage of doing loop. There are two reasons for this: s minimal because the pressure drop between

e and the drum is increased. This reduces the @ The flue gas temperature after the economizer

reducing the amount of heat available for the generated because the saturation temperature heating loop. re evaporator is raised. If the water separation

@ The heat required for feedwater heating increa ive enough, the saturated steam can be sent since a greater flow of feedwater is needed for creased steam production. The preheating loop take up more energy. lphur or sulphur-free fuels, further improve-

becomes possible. When the dewpoint is low Other Parameters t gases can preheat a more or less significant

We will not investigate the effects of the other desi ter in a low temperature economizer. Fig. eters here because they differ only insignificantly fro le burning sulphur-free natural gas. The in a simple single-pressure system. eated far enough in a deaerator so that

the water dewpoint of the exhaust ga-

3.1.3 Two-Pressure System OF). Because this temperature is so low, ce in this case under a vacuum. Following

A single-pressure system with a preheating loop pro arator, all the feedwater is heated in a ter waste heat utilization than a simple single-press Nevertheless, that utilization is neither energeticall

SYSTEM MYOUTS 69

bar

:ondenser

70 COMBINED CYCLE GAS Br STEAM TURBINE POWER PLANTS SYSTEM LAYOUTS

Figure 3-33

Fig. 3-33: Simplified Flow Diagram for a Two-Pressure System for Fuel Flow Diagram for a Two-Pressure System for Sulphur Contain Sulphur

1 Compressor 10 High pressure steam bypass ll Feedwater tanwdeaerator

LO High pressure steam bypass 2 Gas turbine

I2 High pressure feed pump 11 Feedwater tanwdeaerator

3 Flue gas bypass (optional) I2 High pressure feed pump 4 High pressure superheater 13 Condensate pump 13 Condensate pump 5 High pressure evaporator 14 Low pressure feed pump

I5 Low pressure evaporator 14 Low pressure feed pump 6 High pressure economizer 16 Low pressure boiler drum

I5 Low pressure evaporator 7 High pressure boiler drum

17 Low pressure preheater 16 Low pressure boiler drum

8 Steam turbine 17 Low pressure economizer 9 Condenser IH Low pressure steam bypass

72 COMBINED CYCLE GAS & STEAM TURBINE POWER PLANTS SYSTEM UYOUTS 73

low pressure economizer to approximately the saturation te ost of the feedwater preheating is still being accomplished perature of the low pressure steam. It is then admitted to t exhaust gas heat. The boiler feed-water temperature must low pressure drum. Next, a high pressure feedwater pump below the water dewpoint (if the fuel is sulphur-free) culates the feedwater for the high pressure evaporator from id dewpoint (if it contains sulphur). low pressure drum into the high pressure steam generator. this case, too, it is possible to supply the low pressure steam advantage here is the reduced efficiency resulting from

the turbine either as saturated steam or as slightly superheat ing higher quality steam from the turbine. Moreover, densation pressure is low, it may become necessary e another low pressure pre-heater heated with ex-

In addition to this system, there are further variants p team in order to reduce wetness at the end of the tur- Most of these are not as good thermodynamically, but o ould reduce the power output slightly further. tain operational advantages.

f Two-Pressure Combined-Cycle Plants One example is shown in Fig. 3-35, where the high pr ss here examples of two typical two-pressure

and low pressure feed-water are separated directly after th plants, both based on the same gas turbine as water tank. The low pressure economizer shown in Fig. -pressure systems. The first is designed for burn- therefore divided into a low pressure economizer for t nd for burning sulphur-free natural gas. pressure feedwater and a high pressure economizer for t step in preheating the high pressure feedwater. This syst wo-pressure system for fuels containing sulphur the following advantages:

s the main technical data for this unit. The ma-

@ better availability, since the high pressure portio rom the single-pressure system with a preheat- remain in operation even if either the low press 3-stage feedwater preheating. Two pump or the circulating pump fails aters heated with extraction steam reduce

@ fewer problems with steaming out in the low m required for the deaearator, which sup- economizer during part-load operation. mount of excess steam to the low pressure

e it produces additional mechanical energy. On the other hand, a slight reduction of about 5% i ssure has been raised to 60 bar (870 psia) in

sure steam generation must be accepted in most cas e efficiency of the steam process. Unlike the

Another possibility that operates without vacuum ems, this system is not significantly affected

is shown in Fig. 3-36. The deaerator here operates heat utilization in the high pressure portion

overpressure. To do this, it requires extraction steam o oiler because the heat that is not utilized is

quality than that in a system with vacuum deaeration. ow pressure portion. Table 3-3 (page 83)

flows within reasonable limits, the condensate is preh hnical data of this plant.

the feed-water in a water-to-water heat exchanger. Th

SYSTEM LAYOUTS 75

SYSTEM LAYOUTS 79

use its temperature is low, the additional heat absorbed ciently be converted into work. A large portion of off again in the condenser. The exhaust steam flow

urbine and the cooling water system are approximately

the improvement in efficiency is great enough so that

ows the heat flow diagram for Example 2. Com- ple single-pressure system (See Fig. 3-5), the sharp

temperatureheat diagram of the steam gen- s, approximately 70% of the heat exchange high pressure portion and approx. 30% in the

re and low pressure steam generation respec-

g water temperatures, for the plant zer. As in the case of single-pressure air temperature affects overall ef- ly. In this case, the gradient is even

80 COMBINED CYCLE GAS & STEAM TURBINE POWER PLANTS SYSTEM MYOUTS 81

Figure 3-39

EXHAUST GAS

HEAT TRANSFER

Fig. 3-39: Energy Flow Diagram for the Two-Pressure Combined-C Low Pressure Economizer

Q Energy input V1 Condenser Loss V2 Stack Loss V3 Loss due to Radiation in the Waste Heat

1 bar

Steam


s of the two curves explains the contrary functions

to generate high quality steam, that of the the remaining waste heat as fully as possible, mplished only if the pressure in the evapor-

ow. However, there are two reasons why the

flow of steam becomes very large, result- pondingly large duct cross-sections.

the rate of waste heat energy utilization in ction of the low pressure live steam pressure.

here, too, be as high as possible, without

q~~ Eficiency of steam process

P ~ , ~ . ~ ~ h w pressure live steam pressure before turbine vaporator, to be sure, a higher superheat- High temperature live steam temperature 475% High pressure live steam pressure super-heater provides the advant- Low pressure live steam temperature

1, the lack of superheating is com-

e high pressure steam after eam at the mixing point in

SYSTEM LAYOUTS 89

2 4 0 260

the Efficiency c

O C

)f the

92 COMBINED CYCLE G A S & STEAM TURBINE POWER PLANTS

Figure 3-47

Fig. 3-47:

SYSTEM LAYOUTS 93

en the high pressure and the low pressure steam portions. reason, it is about half as great as with single-pressure

where the useful pressure drop in the low pressure steam imately half that of the high pressure steam. For that e pinch point selected for the low pressure portion

so not be too low.

ary: With a two-pressure system, the pinch points of gh pressure and the low pressure evaporators have fect on the efficiency of the steam proces than with

sure systems. If equal economic value is attached to cy, then, the pinch points selected for two-pressure uld be larger than those for single-pressure systems. eration is purely academic, however, since two- ems are selected only where efficiency is valued hat in turn means low pinch points.

ows the relative efficiency of the steam process of the pinch points of the high and low pressure

xhaust Gas Temerature

in the exhaust gas temperature lowers the effi- am process. This reduction, however, is less pro- n with the single-pressure system (Cf., Fig. 3-24) y utilization rate does not drop off as quickly. turbine exhaust gas temperature is, the more re system makes. Fig. 3-49 shows the ratio encies of the two-pressure and the simple

ses as a function of the gas turbine exhaust a theoretical exhaust gas temperature of

is ratio is pracically equal to 1. This fact ems that have supplementary firing (refer

mgle- Gas

96 COMBINED CYCLE GAS & STEAM TURBINE POWER PLAN SYSTEM MYOUTS 97

3.1.4 Special Systems em with steam injection is finding steam at the suit- In addition to the four systems discussed in Section 3. re level. Live steam is generally either at too high

3.1.3, there are also others that can at times prove usef re live steam) or too low (low pressure live steam) will cite two examples: epending on the gas turbine and the load involved,

level required for steam injection is, at least for @ A system with steam or water injection int a1 gas turbines, between 15 and 25 bar (203 and

turbine to reduce nitrogen oxide emissions ( ng reduced high pressure steam is the simplest and @ A system using a single waste heat boiler for t ensive solution but is exergetically undesirable.

turbines ows an improved solution which employs a three-

The system with steam or water injection is taking on er with a standard high pressure portion, a importance as environmental protection regulations be sure evaporator for generating the injection steam, more stringent; the system using a single waste heat sure portion for preheating the feedwater. The two gas turbines is of interest mainly for smaller machi relatively complicated. It can be simplified by unit power ratings of between 5 and 30 MW. m from the turbine (Fig. 3-51), which makes it

the standard systems without additional equip- System with Steam or Water Injection into the Gas Turbine

Environmental protection laws such as those curre to which of these two systems is the better must fect in the USA, Japan, and in most European countri one case to the next. It is certain that the three- that the NOx levels in the exhaust be very low. With p ment attains a slightly higher efficiency at full gas turbine combustors, special measures must be t der to maintain these levels.

e disadvantage of a solution employing steam One way to reduce the formation of NOx during urbine might be part-load operation in instal-

is by lowering the temperature of the flame, sine ral gas turbines. Unless all the gas turbines are of the reaction producing NOx is noticeably rapid ressure at the extraction point decreases so high temperatures. Injecting water or steam into th st cases inadequate. It thus becomes neces- can produce the temperature reduction desired (ref to live steam, which again negatively affects 9.1). e-pressure system is better in this regard.

f the amount of injection steam generated For gas turbines alone, with no waste heat boile enough live steam need be used to cover

to inject water but efficiency is lower than with ste

mparison between the single-pressure sys- loop without steam injection into the gas system with injection of extracted steam

100 COMBINED CYCLE GAS & STEAM TURBINE POWER PUIN SYSTEM UIYOUTS 101

or with water injection. This comparison, of course, do sharp drop in efficiency, approx. 2% with steam and al- in itself possess any absolute validity because the amou % with water injection explains why all gas turbine man- steam injected varies with specifications and type of gas tu rers are working on the development of dry low NOx It does, however, show trends. es as presented in Section 10.3 in order to attain low

sion levels without requiring water or steam injection. The comparison shows that the system with steam inj

has a slightly higher overall power output and less w possible to conceive of a solution that no longer has to be dissipated in the condenser than the dry system. rbine at all: All of the steam is directed into the fact indicates that there is less waste heat from the turbin 91. Fig. 3-52 shows one such system which could the condenser, which also makes the installation less expe as a peaking unit in countries where water is plen- Dis-advantages, on the other hand, are its lower effic e and attains an efficiency higher than that of the greater amount of additional water required. In so alone. However, if the steam flow injected is then, the solution with steam injection can be more e e than approx. 2 - 4% of the air mass flow, major that the standard solution. The prerequisite for this, s must be made to the gas turbine, principally mo- is having available a low cost source of additional w he compressor. This system is therefore only of system with water injection has the lowest efficiency b interest. Because its efficiency is lower and its put is approx. 7% greater. Water consumption is less ion far greater than with the normal combined because the water has a better cooling effect than st for its economical application is quite limited.

Table 3-5: Comparison of the single-pressure system highly sophisticated systems are being marketed with preheater loop with and without st r such names as STIG (Steam-Injected Gas Tur- or water injection, Fuel Oil #2 etc. [45], [46]. They all suffer from the disad-

er consumption is high, and that their efficiency

Injection in the gas turbine Water Steam Dry ot as high as in normal combined-cycle plants. ire specially designed gas turbines, which g their acceptance. STIG systems are

Gas turbine output 73 800 76 000 68 400 k r smaller cogeneration plants with aero- Steam turbine output 38 600 31 200 36 800 k . Fig. 3-53 shows one such system, with

255 700 239 500 228 000 k for two gas turbines. Station service power 1 300 1 200 1 200 k Net power output 111 100 106 500 104 000

d to simplify and reduce the costs of the Waste heat in condenser 78 900 61 800 76 100 k r the combined-cycle plant includes sev-

interest mainly with smaller gas tur- s possible with this system whenever

r is provided to serve two gas turbines.

102 C O M B I N E D CYCLE GAS & STEAM TURBINE POWER PLANTS SYSTEM LAYOUTS

The cost reduction is less with larger machines. The advantage of this solution (Fig. 3-54) are:

@ savings for the evaporator

@ simpler steam circuit

required, a cost increase results which cancels out to a

The reduction in availability can be considered mod the unfired waste heat boiler is a reliable component

In summary, it can be stated that this arrangement esting only for combined-cycle plants employing small bines that one intends to equip with a flue gas bypass

exhaust gas channel, since these can be built without pensive flue gas ducts. The boiler can be placed bet

ment without a steam turbine, with 100% steam injection in th

Uf Steam injection ll Feedwater tanwdeaerator

ipment) 12 High pressure feedwater pump In an open-cycle gas turbine, only 25-35% of the 13 Condensate pump

14 Low pressure evaporator 15 Low pressure feedwater pump

used for an supplementary firing in the steam gen 16 L ~ W pressure drum

renders the combined-cycle process even more v regard to design, operation, and choice of fuel.

Earlier combined-cycle installations generally h tary firing. The fact that that is frequently no 1 today can be attributed to progress in the devel

LC gas


gas turbine. Thermodynamic interest in supplementary mbined-cycle Plants with Limited decreases as the gas turbine inlet temperature rises (Secti pplementary Firing Fig. 3-55 (p. 109) shows the efliciency of the combined-c lementary firing heats the exhausts gas to at most cess using the gas turbine inlet temperature as a par OC (1472 to 1672 OF). The arrangement of the steam The curves are valid for single-pressure steam circuits ed is similar to that of installations without supple- supplementary firing. Older gas turbines had low tur temperatures. With these machines, an increase in ring. Up to temperatures of 750 OC (1382 OF), simple ture to 750°C (1382°F) improves overall efficiency. B boilers can be used, without cooling of the combus- point, supplementary firing brings increases only in that point, a cooling similar to that used

tional steam generator is necessary.

In gas turbines with inlet temperatures in excess ed are oil or gas. With a simple waste heat boiler (1832 OF), the gain is negligible even in the lower rang combustion chambers, gas is the best fuel because pressure processes, only a slight gain in efficiency can iation and ease of ignition. with a supplementary firing to 750°C (1382 OF). Co pressure processes, however, attain their maximum ows that the efficiency attains a maximum at a when utilizing the waste heat alone. er the supplementary firing) of 750 O C (1382 OF).

use the heat exchange in the economizer is op- As gas turbine inlet temperatures keep increasing, since the curves for flue gas and water tem-

tance of supplementary firing will diminish even allel. The exchange of heat can therefore take ertheless, the increased operating and fuel flexi mum loss of exergy. Fig. 3-56 (p. 110) shows combined-cycle with supplementary firing may b eat diagrams for temperatures of 500" (932" in special cases. Particularly in installations use and 1000°C (1832°F) after supplementary fir- ation of heat and power, this arrangement make e temperature curves in the economizer are control the electrical and thermal outputs separ he minimum difference in temperature on Section 4). This pattern is the same as that for a waste

supplementary firing (refer to Sect. 1). At Combined-cycle installations with supplementa r hand, the minimum difference in temper-

one of two categories: the water end-is at the inlet to the rn corresponds to hat of a conventional

o units with limited supplementary firing, w similar to units without supplementary firi

o units with maximum supplementary firin most of the oxygen contained in the gas the supplementary firing is the limit case hausts is utilized. This type of power pla nee in temperature along the entire econ- the conventional steam process. t the exhausts can practically be cooled

r temperature, thereby eliminating the evaporators (Sections 3.1.2 and 3.1.3).

108 COMBINED CYCLE GAS & STEAM TURB SYSTEM OllYOUTS 1

Unlike conventional power plants, the feedwater tempe here depends solely on the sulphuric acid dewpoint (Section Thermodynamic improvement by multi-stage preheating higher temperatures serves no purpose. This pattern corr to that of combined-cycle plants without a supplementa waste heat boiler.

Example of a Combined-Cycle Unit with Limited Supplementary Firing

Fig. 3-57 shows the heat balance of a typi plant with supplementary firing to 750 O C (1 the gas turbine is the same 70 MW machine as that us examples without supplementary firing. T sulphur-free natural gas, which produces results opti regard to efficiency. Natural gas has the fu it can be burned easily in a waste heat bo combustion chamber. With oil, even that is (Section 5.2). The basic arrangement for this insta same as that for the purely one-pressure system in 3. the fuel contains no sulphur, the feedwater be reduced to 60 OC (140°F). Deaeration therefor

the refired combined-cycle installation as a function of th under a vacuum. perature after supplementary firing and gas turbine inlet

Fig. 3-58 shows the corresponding Temperature of the combined-cycle plant

One can see the optimum temperature pattern in th resulting in a low stack temperature. This is low pressure evaporator would bring not further i ization of the waste heat energy at full load. At pa ever, or when the supplementary firing is switched temperature rises. For installations that are fr at part loads, it can thus make economic sense rangement with a preheater loop. The sa applies to plants with a temperature after su lower than 750°C (1382 OF) at their design poin

110 COMBlNED CYCLE GAS & STEAM TURBINE POWER

Figure 3-66

a l

Fig. 3-56: Temperaturaeat diagram for 500' (93Z°F), 750" (1382°F) ( (1832°F) (c) after supplementary firing

t Temperature in "C Q/Qk Heat exchanged A Flue gas B WaterISteam

SYSTEM LAYOUTS 711


The Influence of Ambient Conditions natural gas. When burning oil, the paths of the curves for on Power Output and Efficiency single-pressure system would not be significantly changed,

there would be less difference between the single and If the temperature after supplementary firing is constant, t essure systems (Section 3.6).

effects of air pressure and air temperature are similar to th in installations without supplementary firing. These two des hould note the decreasing difference between the effi- parameters have a very pronounced influence on the pow s of single and two-pressure systems as the flue gas tem- put, but the efficiency remains to a large extent unaffe re after supplementary firing increases. The curve in Fig.

ms that the two-pressure process provides no advan- Because the steam turbine is providing a greater po the single-pressure process at temperatures above

the total power output, the temperature of the condens 2 OF). Here, too, the improvement at low flue gas tem- ing medium has a stronger effect on the overall power o is greater with the two-pressure process. It makes no and the overall efficiency. Its effect on the steam process at all for the steam turbine whether this flue gas tem- is similar to that in plants without supplementary firin s attained directly from the gas turbine or by means 3-22). However, because of the higher live steam data, t entary firing. The results indicated in Section 3.1.3 of enthalpy drop is slightly reduced. re also valid for installations with supplementary fir-

The Influence of the Most Important Design Parameters on Power Output and Efficiency ows another reason why the machine behaves in Flue gas Temperature after Supplementary Firi e, the rate of energy utilization in the single-

The temperature after the supplementary firing i continues to rise as the temperature after the

important design parameter because it strongly infl ring increases, up to 750 OC (1382 OF). This is

power output and the design of the plant. sing stack temperature and the increasing tem- tial between the steam generator inlet and out-

Fig. 3-59 shows how relative power output and ate of thermal energy utilization increases as

pend on the temperature after supplementary fi s, the improvement possible with a two- tom limit, 525 OC (977 OF), represents utilization of t omes continually smaller.

waste heat alone. Two different systems have b the diagrams:

for a combined-cycle plant with supple- @ a single-pressure system (Fig. 3-57) 50 OC (1382 OF) are comparable to those @ a two-pressure system (Fig. 3-34) rbine plant with the same power out-

e. The live steam temperatures have The basis for comparison is the single-pr o the pressure, according to values

out supplementary firing. Calculations assu wer plants. As is the case in

l6 COMBINED CYCLE GAS & STEAM TURBINE SYSTEM UYOUTS 117

Figure 3-59

firing is practically unaffected by the live reduction in exergy losses at higher live

re directly affects overall efficiency of the

on in the steam generator, the tempera-

kes it necessary to preheat to a higher eating should take place in several stages.

be employed to reach the required 110

the higher feedwater temperature.

an 750 OC (1382 O F ) , a system with a pre- asonable way to make optimum use of . The more often the unit is run at part

ombined-cycle plants, the increased

Fig. 3-59: Effect of the Temperature after Supplement output being provided by the steam Efficiency of the Combined-Cycle Plant and Utilization

ser pressure is of more importance nts which merely utilize the waste

P ~ / P R ~ ~ Relative Power output of the combined-cycle plant

VK/VREF Relative efficiency of the combined-cycle plant WB Rate of utilization of waste heat energy kF Flue gas temperature after supplementary firing ative effect of a change in condenser * Reference = Plant without supplementary firing (52

118 COMBINED CYCLE GAS & STEAM TURBINE POWER SYSTEM LAYOUTS 119

effect it has on the power output of the steam

cycle plants have a greater exhaust steam flo sing only part of the oxygen remaining in

put, since there is less preheating of the feed- ine and a steam turbine with reheat. In a reason, the condenser pressure exerts a greater i pe, the regenerative air pre-heater usu- power output from the steam turbine.

flue gas is being used as the combustion 3.2.2 Combined-Cycle Plants with Max

Supplementary Firing w to flue gas flow is much greater than ue to considerations of exergy only a

The basic idea for combined-cycle plants wit should be directed through the econo- ited supplementary firing in the waste heat bo d with flue gas. The rest flows as nor- the best possible use of the gas turbine's wast

rt-flow economizer is therefore ideal not the gas turbine but the conventional steam water and steam temperatures run was to provide a prior gas turbine in order to i

-free natural gas, the energy is the proved air blower with an air heater built in. to provide electricity. pressure portion of the feed water

nt is advantageous, however, only if This approach is reflected in the ratio of t usts is very low (Fig. 3-62). steam and the gas turbines. Depending on the

air, this is between 4 and 10, as compared to cycle plants without supplementary firing.

eases the availability of the a conventional steam power

cy in operation of the fan is a unit with steam reheating a water heating.

eat importance here.

The number of possible systems av tically all known steam processes ca

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SYSTEM LAYOUTS 129

steam power plant because of thermody- Thus, in most cases, no extraction points

the steam turbine, which increases the turbine. However, the generator and

e able to handle this additional power cases that can lead to restrictions.

ignificant gain in efficiency that can be ng. The power output of the plant as n tripled. A supplementary firing could output because the gas turbine exhausts

vering a portion of the heat demand of

128 COMBINED CYCLE GAS & STEAM TURBINE Po

Figure 3-64

Fig. 3-64: Combined-Cycle Plant with Existing Steam Turbine

1 Gas Turbine 9 Condenser 2 Compressor 10 Steam bypass 3 Flue gas bypass ll Feedwater tanwdeaerator 4 Superheater 12 HP Feedwater pump 5 Evaporator 13 6 Economizer 14 LP evaporator 7 Drum 25 LP Feedwater pump 8 Steam Turbine 16 LP drum

1 700 1 200 kW

107 000 228 000 kW

plementary firing in repowering is the tation to the live steam data of the orig-

ign of the gas turbine has been stan- its on the amount of steam that can

o not necessarily lie close to the design turbine.

130 COMBINED CYCLE GAS & STEAM TURBINE PO SYSTEM LAYOUTS 737

Another approach to repowering can be used cle ~nstallations with dern steam Power plants equipped with reheat Gas Turbines [63 to 681 To improve the efficiency of such a unit, the fie be replaced in supplying the oxygen needed for c losed-cycle process is, however, much less installing a new gas turbine before the existing tor. Here, the existing steam boiler continues in sisting of compressor, heater, turbine, be adapted to its new operating mode. The modific ly, the medium can be any gas1 but al- are due mainly to the much higher temperat at have been built employ air. Helium bine exhausts (approx. 500 OC/ 932 OF as camp ossible media, with helium in particular 572 - 662 OF for the fresh air after an air preh nuclear power stations.

Parts that must be modified are the: gas turbine can be raised by employing

@ burners ems, e.g., a recuperator, compressor @ fresh air ducts ng, or the like. However, as was the

@ perhaps the reheater as turbines, the simpler arrangements

A waste heat recovery system must be installe generator to handle most of the condensate and f the closed-cycle gas turbine is the great

ing. The complete plant appears similar to tho lection of the fuel. In addition to oil or 3-60 and 3-62. 1 can also be used. Its main disadvantage

lied to the process via a heat exchanger, The major problems that arise with this typ e inlet temperature to levels bwer than

are in connection with: turbines. According to [64], efficiencies Id be attainable. Fig. 3-66 shows the way

a) space availability for installation of the g of the combined-cycle process depends and the waste heat recovery system ssure ratio and the gas turbine inlet tern-

b) adaptation of the boiler and the overall cept to the new mode of operation

tial application involves nuclear Power For steam Power plants that burn gas or oil ed reactors (high temperature reactors

one very interesting possibility for raising s). In these plants, the gas would be in than 10% and power output by 20 to 30% a reactor. The maximum process temper- vestment Costs. With coal-burning units, ther the highest process temperatures attain- economic gain because the conversion itself is there is less improvement in efficiency.

today at 950 O C (1742 OF).

134 COMBINED CYCLE GAS & STEAM TURBINE POWER SYSTEM L4YOUTS 135

The use of combined-cycle plants with closed- nt advantage is the small, compact steam gen- bines can be considered for the following applic e and the greater speed

r is better than in a con- @ burning coal in a fluidized bed rator. Consequently, the surface required

high temperature reactors. hat theoretically should

Fig. 3-67 shows one possible arrangement with combustion, a closed-cycle air or helium the second type. Here, un- sequent reheating steam process. A com oying a simple charging group, advant- a high temperature reactor, helium turbine, a gas inlet temperature. The steam can be seen in Fig. 3-68. In neither case can a aces the gas turbine combustor. Net effi- be expected in the near future because are within the range of possibility. That economic hurdles are too great. have no thermodynamic

mbined-cycle plants with 3.5 Pressurized Steam Generators even these values are

by simple combined-cycle plants with- As one final combination we should mention a g which today exceed 50%. The ques- a pressurized steam generator. This type of' pow

to whether there will be any economic fall into one of two categories: gas-fired plant of this type in the future.

0 installations with a simple charging group, r efficiency, it has the following further or may not provide (a small amount of) ele power mbined-cycle plant with

0 installations with gas turbines and subse mizers

uiring a special steam generator Fig. 3-69 shows the diagram of the principl

first of these arrangements. Power plants lik n of the gas turbine and steam quite early: the more than 100 Velox boil veri operate in this way. However, such ins real future (at least for oil or gas-burning pla cannot provide any genuine thermodynamic the following advantages: conventional steam plants. The gas turbi a very low temperature level because it is or- though this advantage is from the steam generator as a working m ut by the subsequent economizer is being made of the high temperature pot machine, and its power output is for that re process possible because Typical efficiencies of such units are in th of the steam generator.

SYSTEM LQYOUTS 137

140 COMBINED CYCLE GAS & STEAM TURBINE POWER P u

The balance is more likely to be on the negative sid oil or gas-fired plant.

Plants with pressurized steam generators could ne still become very interesting in the future because t a possibility of burning coal cleanly in a pressurized bed combustor (PFBC). Refer to Section 10 below for information about such plants.

3.6 Summary and Evaluation of the Vario Arrangements Possible

In our evaluation of the various arrangements p will omit from consideration systems with closed-cy bines or pressurized steam generators because they of academic interest only. The following six exarnp compared below:

0 the single-pressure sytem (Fig. 3-4)

the single-pressure system with a preheater 3-27)

@ the two-pressure system for fuels containin (Fig. 3-37)

@ the two-pressure system for fuels containin phur (Fig. 3-38)

@ the combined-cycle plant with limited suppl firing (Fig. 3-57)

@ the combined-cycle plant with maximum s tary firing (Fig. 3-60)

All arrangements are based on the same gas t approx. 70 MW and are as a result directly compa

Gas Coal

ST Steam turbine Suppl. Supplementary

-fired steam generator with a low pressure part-flow econo

point is the high net efficiency of the gas- ystem. The unit with limited supplementary

r, far behind: its efficiency is less only by Its power output, however, is about 40% am turbine produces about twice as much ent can be of interest whenever a higher 'nable with utilization of the waste heat reover, the specific investment costs re- to be lower than those for the two-pres-

142 COMBINED CYCLE GAS & STEAM TURBINE POWER P

sure system. But this type of plant is more complex become less and less attractive in the future as gas tur temperatures continue to increase.

Efficiency is only one of the important criteria wh on the selection of a power plant. A second is the is difficult in a Handbook of this type to provide ex Only comparative prices are possible. The basis fo is in all cases the simple single-pressure system a ative prices are valid as specific prices for install comparable power rating (Table 3-10).

Table 3-10: Comparison of Specific Prices for the

SYSTEM LAYOUTS 143

nt consideration is the amount of cooling wa-

arison of the Amount of Cooling Water The net efficiency of the combined-cycle with m

plementary firing is poor, but it can cover 70% o quirements with coal, which is frequently an adva

133.4 670

3180 15940

109.0 145.3 348.8

21.9 46.5

supplementary firing need be- er installed MW because in these

ional low exergy heat is being supplied m the exhaust gas. Because this energy into mechanical energy, it must in large again in the condenser.

unt of cooling water required, the maximum supplementary firing turns

Various Systems, in %

Single- Single- 2-Pressure 2-Pressure Limited Pressure Pressure, Fuel with Fuel with Suppl.

Preheater Sulphur no Sulphur Firing Loop

Relative price 100" 101-103 105-108 106-110 103-110

* Basis for comparison

The higher specific price for the unit with a c rating when employing maximum supplemen to the fact that a steam plant is more expensi bine, which means that the relative price for a proportionally large steam component will This is especially true for a plant with a coal-bu today usually requires installation of a scrubb sulphur from the exhaust gases.