Thermodynamic analysis of a coal-based combined cycle power plant

Pergmnon

Hint Recovt.ry S,~mmu & CliP Vol. 15, No. 2, pp. 115-129, 1995 ~ h t © 1994. msevi~ ~ Lid W~.4.~94)00~-7 ~ m Great n r i t ~ . AU ri~m ~ m ~ 0890-4332/95 $9..q0 + .00

THERMODYNAMIC ANALYSIS OF A COAL-BASED COMBINED CYCLE POWER PLANT P. K. Nag and D. Raha Department of Mechanical Engineering Indian Institute of Technology Kharagpur 721 302, India ABSTRACT A thermodynamic analysis of a combined cycle power plant using pressurized circulating fluidized beds for partial gasification and combustion of coal has been made on the basis of both first law and second law. The Redlich-Kwong equation of state is used for evaluation of properties of air at high pressures in the topping gas turbine plant. A dual pressure steam cycle is considered in the bottoming plant for reducing irreversibility in heat transfer from gas to water and steam. The effects of pressure ratio and peak cycle temperature ratio of the gas cycle and the lower saturation pressure of the steam cycle on the overall performance of the combined plant have been evaluated. INTRODUCTION Combined cycle power plants recently received worldwide interest not only because of their high energy conversion efficiency and low emission capability, but also due to their potential of using coal, the most widespread abundantly available fossil fuel. In a typical coal based combined cycle coal is gasified, either partly or wholly, and the synthetic gas produced is supplied directly to the combustion chamber of the gas turbine. The exhaust from the gas turbine is used to generate steam which drives a conventional bottoming steam plant. In an integrated gasification combined cycle (IGCC), coal is gasified completely. Here the working fluids work over a large temperature range, say from 1100° C to 550° C in the gas turbine and 550° C to the ambient temperature in the steam turbine, thus achieving an overall efficiency approaching 50% [1]. Reduced emission levels, lower investment cost, higher availability, lower quality fuels, and shorter design and construction periods are other attractive features of a combined cycle power plant. Horlock [2] presented detailed thermodynamic analyses of various schemes of combined power plants. An exergy analysis of combined cycles was given by E1-Masri [3,4] in two parts: the first part dealing with the air-cooled Brayton cycle gas turbines treating air as an ideal gas with constant properties, and the second part involving the analysis and optimization of dual pressure steam bottoming cycles. Cerri [5] gave a parametric analysis of a gas-steam plant determining the independent parameters of the cycle and their limitations, and also carried out a cycle optimization based on first law. Partial gasification of coal and subsequent combustion of the char produced in pressurized circulating fluidized beds (PCFB) have been proposed in relation to a combined cycle plant with the potential of achieving still higher plant efficiency and lower environmental pollution. ~RS lS:2-c

115

116

P.K. Nag and D. Raha

Presently no thermodynamic analysis of such a cycle is available. A thermodynamic analysis of the partial gasification plant, both on first and second laws, has been carried out in this paper. The Redlich-Kwong equation of state was considered for the evaluation of properties of air in the topping gas turbine plant. The effects of certain pertinent variables on the plant performance were evaluated. THERMODYNAMIC MODEL A schematic of the partial gasification combustion power plant for which analysis has been made is shown in Figure 1. The air supplied by the gas turbine (GT) compressor is split into two parts - air from the low pressure (LP) compressor goes to the GT combustor and that from the high pressure (HP) compressor goes to the gasifier and the char combustor. The coal is pyrolyzed/carbonized in the PCFB gasifier and the char produced is fed into the atmospheric CFB combustor unit. The low heating value gas from the pyrolyzer is burnt in the GT combustor after passing through a hot gas cleanup system. A dual pressure steam cycle is selected to reduce irreversibility in heat transfer from the gas to water/steam [2]. The residual energy in the gas exhausting from the GT is recovered in the heat recovery steam generator (HRSG). This is used for raising steam in the low pressure cycle. The energy released by combustion of char in the CFB combustor is used to raise steam for operating in the high pressure cycle. The model is later modified to make the char combustor also act under pressurized condition, ie., as a PCFB unit. GAS TURBINE PLANT The working fluid for the GT plant is chosen as air for which Redlich-Kwong equation of state has been used to evaluate properties at pressures higher than atmospheric and ideal gas equation of state at atmospheric pressure. Redlich-Kwongequation of state, as given below, is one of the simpler equations with only two specific constants, but it has considerable accuracy over a wide range of p-v-T conditions[7]: RT a P - v - b " T 1/2v(v + b)

(1)

where the constants in terms of critical data are 2.5 c = 0.42748 R2 e"c and b = 0.08664 R "ecT-~ T

a

For evaluating thermodynamic properties, like enthalpy and entropy change in various processes, it is necessary to determine the specific volume at a given state specified by temperature and pressure. But the determination of specific volume becomes complicated in the case of Redlich-Kwong equation of state because it yields a third order polynomial in terms of v. So an iteration scheme, Newton-Raphson method, has been used to solve for v by writing RT a f(v) = p - v - b + T1/2v(v+b) •

Analysis of a coal-based power plant

1! 7

and f~(v) = - R.T In (v - b) +

V

In v + b

(2)

So, the iteration scheme becomes

ffnk)

Vk+ 1 = v k - ~(nk)

(3)

A computer procedure has been developed to compute v by the iteration technique as mentioned earlier. To evaluate gas temperatures T 2 and T 4 after isentropic compression and expansion respectively, the compression process 1-2, for example, is split into three processes as shown in Figure 2. By using gedlich-Kwong equation

(sa-sl) +(s b-sa) +(s 2-sb) =0 n a -b [R. In n 1 "b "

(na+ b) a na 1/---'~-In 2T 1 b (nl + b'] ) nl

+ g[a In T-' + T21

b (T2 "T1 ' + 2g(T22 - T21) + q3(T23 - T31) + e4(T24 "T41)]

n2 -b + [R In nb. b "

(n2+ b) a n2 1 =0 3/----7In 2T 2 b (nh + b) j nb

(4)

T2 where

Sb-S a = J T - q ~ , C p = a + b T + g T 2 + q T

3 +eT 4

T1 Equation (4) is used to evaluate T2. The T4 can also be calculated in a similar way. To evaluate the enthalpy changes in the various processes the following results are obtained:

(nz + b) 1.5a r na 1 h a - h 1 =(Pana - Plnl) - ~1/2 ) In T1 b t(nl + b) J nl hb - h a = R[a(t 2 - t 1) + ~b ((T 22 - T~)+3g(T32 - W31 ) + 40 (T24 _ T41) + e-5 (T~-T51 )

118

P.K. Nag and D. Raha

(nT. + b)

1.5a r n2 h2 - hb = (p2n2 - Pbnb) - 1/2- ) in ,(nb + b) ] T2 b nb

(5)

By summing up the above three equations, the specific work of compression, (h2-hl), is obtained. In the same way, the specific work output of the turbine, (h3 "h4), is also evaluated. The polytropic efficiency of the compressor and turbine is each assumed to be 0.85. GASIFIER

Since the gasifier is pressurized and air blown, a Lurgi gasifier has been considered for analysis. The volumetric composition of the flue gas typical of a Lurgi gasifier as given below by Perry [8] is taken constant for all conditions of operations. CO 2 13.5 %, CO 16.5 %, H 2 23.8 %, CH4 4.0 %, N 2 41.3 % , others 0.9 % (by volume). The gasifier operates between 8 to 30 bar giving a gas exit temperature of 500°C to 700°C. The coal fed to the gasifier has been assumed to be of the following composition: Proximate analysis: Fixed carbon 42.6 %, Volatile matter 35.4 % , Moisture 8.8 % and Ash 13.2 % (by mass). Ultimate analysis: Carbon 62.2 %, Hydrogen 5.2 %, Oxygen 14.3 %, Sulphur 3.8 %, Nitrogen 1.3 %, and Ash 13.2 % (by mass, on dry basis). In 100 kg of coal, the fixed carbon is 42.6 kg. The mass of carbon present in the volatile matter is (62.2 - 42.6), i.e., 19.6 kg. It is assumed that the whole mass of carbon in volatile matter and 10 % of fixed carbon, (19.6 + 4.26), or 23.86 kg get gasified,while the remainder (42.6 - 4.26) ,i.e., 38.34 kg escapes as char and is combusted in the CFB combustor. The following simplified equation of the reaction takes place in the gasifier: (5.18 C + 1.47 H 2 + 0.05 N~ + 0.45 02 + 0.12 S) coal +

0.63 ( 02 + 3.76 N2) + 3.2 H~O air steam

= .5.82

0.135 CO2 + 0.165 CO + 0.238 H 2 + 0.04CH4 + 0.413 N2 fuel gas

2.82 H~O + 3.2C + Ash + S steam char Therefore,

(6)

Analysis o f a coal-based power plant

119

coal: air: steam = 1 0 0 : 8 6 . 8 : 57.6 Fuel gas = 135.25 + 50.76 = 186 kg/100 kg coal Char = 38.4 + 13.2 + 3.8 = 55.4 kg / 100 kg coal It is assumed that (a) steam enters at saturation temperature corresponding to the pressure of the incoming air, (b) only a part of hydrogen enters into gasification, the rest escaping with char and ash, (c) the coal is fed at T 0, P0, and (d) there is no pressure drop in the gasifier. The heating value of coal is obtained from stoichiometric quantity of air being used for complete combtistion of unit mass of coal yielding the following reaction (5.18C + 2.6 H 2 + 0.05 N 2 + 0.45 0 2 + 0.12S) + 6.02 ( 0 2 + 3.76 N 2 ) =

5.18 CO 2 + 2.58 H 2 0

+

22.68 N 2 + 0.12 SO 2

H H V o f coal = H p - H g

= 5.18(-393520) + 2.58 (-241820) = 26.6 MJ/kg coal (neglecting energy released by the oxidation of sulphur) To determine the heating value of fuel gas going to the GT combustor and yielding the following reaction 5.82(0.165 CO + 0.238 H 2 + 0.04 CH4) + 1.63 ( 0 2 + 3.76 N2) = 1.193 CO 2 + 1.85 H20 +6.15 N 2 HHV o f f u d gas = =

Hp-H R 1.193 (-393520)+ 1.85 (-241820) -

[5.82{(0.165 (-110530) + 0.04 (-74850)}]

7.932 MJ/kg coal Similarly, HHV of char = 3.2 (-393520) =12.59 MJ/kg coal.

FIRST LAW ANALYSIS The energy balance of each stage of the cycle can be written as Compressor:

hentry + Wc = hair 1 + hair 2

120

P. K, Nag and D. Raha

Gasifier:

hair2 + hcoal + hsteam = hfuel gas + hchar + Qloss

GT combustor: hfuel gas + hairl = hprod gas + Qloss Gas Turbine:

hprod gas = Wt + hgas steam + hstack

Steam cycle:

hchar + hgas steam = Wst + Qcond + Qloss Wst = (mH + mL ) (h3 -h4) + mH ( hl - h 2 ) Qcond = (mH + mL ) ( h4 - h5)

On the overall basis, hcoal

= (Wt + Wc) + Wst + Qcond" hsteam + (hstac k - hentry) + E losses = Wg + Wst + Qcond + hnet stack + E losses

The energy losses are assumed to be 5% of the energy released in the gasifier as well as the char combustor, and 2.5% in the GT combustor. The condition of steam at inlet to the HP turbine is taken as 120 bar, 400°C and the condenser pressure is assumed to be 0.085 bar. Both the HP and LP turbines are assumed to have an isentropic efficiency of 85%, while the stack temperature is taken as 420 K. The energy balance of the whole system helps to trace how the energy in the primary source, the coal, is used up at different stages of the plant.

SECOND LAW ANALYSIS The chemical availability or exergy of a fuel is the maximum theoretical work obtainable by allowing the fuel to react with oxygen from the environment to produce environmental components of carbon dioxide and water vapor. For a reaction of the type b b Call b + (a + ~) 0 2 = aCO 2 + ~ H20 it is given by [7] b (yt3,,)a+b/4 h a ch = ~f + (a + -,-~)5 0 2 - a(~,)CO 2 - ~(~)H20 + I~T ln(YCO2)mYH20)b/2 where g is the molar Gibbs function and y is the mole fraction of the gas components in the environment. On a unit mass basis the total availability (flow) is given by af = (h - ho) - T O (s - so) + ach where the underlined portion is the thermomechanical contribution and ach is the chemical contribution. The terms ho and so refer to enthalpy and entropy of the system when it is in the dead state. When a difference in availability between states of the same composition is evaluated, the chemical contribution cancels, leaving just the thermomechanical contributions. This will be the case while doing availability analyses for compressors and turbines. However, the chemical contribution will come into picture during analyses of the gasifier, fuel gas combustor and char combustor.

Analysis of a coal-based power plant

121

The chemical availability of the fuel is now traced as it cascades through the cycle, portion of it being destroyed by various components and processes and the balance emerging as shaft work. The availability balances of each component along with irreversibilities are given below: Compressor:

W c = aairl + aair2 + icom p

GT Combustor:

afuel gas + aairl = aprod gas + iGT comb

Gas Turbine:

aprod gas = Wt + agas-steam + igt ch acoal + aair2 + astea m

Gasifier:

= afuel gas + achar + igasifier HRSG:

agas_steam + ast,7 = astac k + ast,3 + ihrsg Char Combustor: achar + ast,8 + aair2 = astl + ichar comb liP Turbine: ast,1 = ast,2 + Wst,1 + ihp, st LP Turbine: ast,4 = ast,5 + Wst,2 + ilp,st Condenser:

ast,5 = ast,6 + icond

On the overall basis, ch

acoal = (Wt

"We) + (Wst, 1 + Wst,2) + X Irrev.

= Wg t + Wst + Y. Irrev.

First law of efficiency of combined cycle W ~ + W.~t =~ of coal Second law of efficiency or effectiveness of combined cycle W~t + W~t = Chemical ~ailability of coal ch The chemical availability of coal, acoal = 5.18(410541) + 1.47(23521 l) + 0.05(691.09) + 0.45(123.32 x 32) = 24.74 MJ/kg coal The availability of char, achar =3.2 x410541.51 = 13.13 MJ/kg The availability Of fuel gas, afuel gas = ath + ach

122

P.K. Nag and D. Raha

= Ey i [h - ho) - To(s - So) + RT o In Yi (Y~) The exergy of a particular fluid at a pertinent location is evaluated and the irreversibility in a process is ascertained from

or,

i =

availability lost by the hot fluid availability gained by the cold fluid

i =

maximum work obtainable - actual work

A computer program has been developed to solve for the efficiency and effectiveness for the whole cycle as well as for individual components and processes. The availability has been nondimensionalized by dividing with coal availability. RESULTS AND DISCUSSION Figure 3 gives the availability balance with the variation of pressure ratio. It is seen that with the increase in pressure ratio the gas cycle output or effectiveness increases, but the steam cycle output decreases. This is because as pressure ratio is increased the gas turbine exit temperature decreases, for the same maximum gas inlet temperature to the turbine, which further reduces the energy and hence exergy recovered in the HRSG. In the limit when the pressure ratio approaches unity, the gas cycle will simply become a burner for the steam cycle. The exergy losses in the gas cycle as well as in the gasifier increase with pressure ratio, whereas those in the steam cycle decrease. Availability or exergy balance with respect to cycle temperature ratio is shown in Figure 4. It is expected that with the increase in peak temperature both gas and steam cycle outputs should increase, and this has been obtained. The gasifier irreversibility does not depend on cycle temperature ratio, and so the irreversibility remains the same. As the cycle temperature is increased less excess air is required and so irreversibilities in the gas cycle, i.e., in the compressor,combustor and turbine, decrease. The steam cycle output as well as the exergy losses in it increase with the increase in peak gas temperature due to the increase in the gas turbine exhaust temperature for the same pressure ratio which improves the energy recovery in the steam cycle and causes more exergy losses during steam generation. Figures 5 and 6 give energy balancing with respect to pressure and temperature ratios respectively. The energy losses at various stages were assumed beforehand. When compared to exergy losses~ interesting conclusions can be drawn. It shows that the losses in combustion chamber and gasifier are very insignificant, whereas the irreversibilities occurring there are quite high. The energy loss in the gasifier is independent of both pressure ratio and peak temperature ratio. The stack loss increases with pressure ratio, but decreases with temperature ratio. Thermal optimization of steam cycle working on gas turbine exhaust differs from that for directly fired boilers in that the value of exergy of gas at inlet to HRSG is less, and significant irreversibilities occur during heat transfer. These losses can be reduced by increasing the number of evaporator pressures. For this reason the dual pressure cycle was selected. The effect of lower evaporating pressure on the steam cycle is shown in Figure 7. The condenser

Analysis of a coal-based power plant

123

and HRSG irreversibilities remain unaffected, whereas the maximum work output increases and the char combustor loss decreases with the increase of this pressure. The results mentioned above were obtained with the char combustor operating as an atmospheric CFBC. The program was then suitably modified to make the same analysis assuming the char combustor as a PCFB unit. The effects of pressure ratio on the efficiencies of the steam and gas cycles for ACFB and PCFB combustors are shown in Figures 8 and 9 respectively. It is seen that the steam cycle efficiency improves due to the use of PCFB whereas the gas cycle efficiency decreases. The extra compressor work required is large enough to bring down the gas cycle efficiency. But the energy recovery in the bottoming steam cycle increases, which causes its efficiency to improve. CONCLUSION A thermodynamic analysis of a coal-based combined cycle power plant from the viewpoints of both first law and second law has been made,

The effects of pressure ratio and peak

temperature ratio of the topping gas cycle as well as those of lower saturation pressure of the bottoming steam cycle on some important performance parameters of the overall combined cycle have been studied. ACKNOWLEDGEMENT

A part of the work was carried out at the Technical University of Nova Scoti under collaborative project funded by the Institutional Linkage Project of the Canadian International Development Agency. The assistance of Prof Prabir Basu in the preparation of this manuscript is gratefully acknowledged.

NOMENCLATURE

a,b a

Cp g h H i P Q

rp R S

T V

constants availability or exergy, kJ/kg mol specific heat, kJ/kgK molar Gibbs function, kJ/kg mol enthalpy, kJ/kg enthalpy of formation, kJ/kg mol irreversibility, kJ/kg mol pressure, kN/M 2

heat loss, kJ pressure ratio, p2/Pl gas constant, kJ/kg K entropy, kJ/kgK absolute temperature, K specific volume, m 3/kg

124

P.K. Nag and D. Raha

W y

work output or input, kJ mole fraction

0

temperature ratio, T3/T 1

Subscript c

critical state

REFERENCES 1. Becker,B. and Schetter, B.,"Gas Turbines Above 150MW for Integrated Coal Gasification Combined Cycles (IGCC)" Trans. ASME, Journal of Engineering for Gas Tirbines and Power, volume 114, p. 660-664 October 1992. 2. Horlock, J.H.,"Combined Power Plants", Pergamon Press, Oxford, 1992. 3. EI-Masri, M.A., "Exergy Analysis of Combined Cycles: Part 1- Air-Cooled Brayton-Cycle Gas Turbines" Trans ASME, Journal of Engineering for Gas Turbines and Power, volume 109 p. 228-236, April 1987 4. Chin, W.W. and EI-Masri, MA., "Exergy Analysis of Combined Cycles", Trans. ASME, Journal of Engineering for Gas Turbines and Power, volume 109 p. 237-243, April, 1987 5. Cerri, G.,"Parametric Analysis of Combined Gas-Steam Cycles", Trans. ASME, Journal of Engineering for Gas Turbines and Power, volume 109, p. 46-54, January, 1987. 6. Basu, P., Ngo, T. and Prasad, B.V.S.S.S.," Hydrodynamics of Pressurized Circulating Fluidized Beds for Partial Gasification and Combustion" Proc. 12th International Conference on Fluidized Bed Combustion, Ed. L. N. Rubow, San Diego, California, May 9-13, 1993, p. 353-360. 7. Moran, MJ. and Shapiro, H., "An Introduction to Engineering Thermodynamics", John Wiley, 1988. 8. Perry, R.H. and Green, D., " Perry's Chemical Engineers' Handbook", Sixth edition, p. 9-23, McGraw-Hill, 1984.

125

Analysis of a coal-based power plant

TO STACK

Fig.1 Model of the combined cycle power plant and the T-S diagram

126

P . K . Nag and D. Raha

T

2 i l---,,.-Jo 1 s

s

(~)

(b)

Fig. 2.(a) Eirayton cycle, ( b) splitting the isentropic compression process 1-2 into three processes l - a , a-b and b - 2 .

1.0

~> "1-

"-r"

0.8

c3 ~J

"~ 0.6 c o

~ . = /

~- 0.4 -

Gosifier irreversibility //--Steam cycle irreversibility / - - Gas cycle irreverlibility //-Steam ~cle output / / /--Gas cycle output

~x 0.2

L4J

J I I l 32 12 16 20 2/, 28 L Pressure ratio, rp Fig. 3 Effect of pressure ratio on exergy values of different cycle parameters 0 0

4

8

Analysis of a coa(-I~,ed power plant

127

1,0

:z: ,1- 0.8 C3

o

0.6 e'-

Steam cycte irreversiblity ,

•

=

•

& 0.4,

g x 0.2

_

L

E .~__._~ .

0

.

.

.

•

,r

' ' . . . . . . .

1~'-Ges cycte output I s Temperature r~fio~ B

Fig. 4 E f f e c t of p e a k t e m p e r a t u r e different

. . . . OQS CyCle !rreversioi,~i~y

~' /-Gasi!ier irreversibility

.

" 1 4

.

Steam cycte output . . //-

r a t i o on e x e r g y

values

of

cycte p a r a m e t e r s

10

~> -r- 0.8 0

~0~

, ..

o

F

Steam c¥cte toss

~0.~

fSteom

_

~ -

cycle output

i/o-<,<'.,,',,'./~"'-. . ....

~ _~ .

. .

_~- Gasifier toss T 0 4

II

"1 12

---.-

I" 16

'

[

~

20

_r 24

I 28

.....

Pressure ratio, rp

Fig. 5 Effect of pressure ratio on different cycle parameters

} 32

! 28

P.K. Nag and D, Raha

1.0

-

;:> -r-~0.8 C3 O ~a

0.6

F

Steam cycle loss

f

Steam cycle output

~q

c~ 0./,

$

/ / ~ $ t o c k loss

0.2

Gas cycle output /(-- Gosifier loss I

1.. "

.

S - - ~ Temperafure rofio,

'

J

6

9

Fig.6 Effect of peak temperature ratio on different cycle pammters

1.0

..t-- 0.8 .1-

Exhaust availability = 0.033

o "6 0.6

~.....~

Char combustor irreversibility

._,__,_.~

Steam cycle output

c

E

0,4

o~ I:l

0.2 Condenser irreversibility HRSG irreversibility

f o

0

I

L

8

16

L

,I

I

24

32

f•

.. J

40

I

t,8

--

i

i

56

6/.

Lower safurafion pressure, bar

Fig. 7 Effect of lower evaporating pressure on steam cycle perometers

Analysisof a coal-basedpower phmt

129

0.26 f eOJ t.I q..

PCFB

0.24

~0.22 I.I

~

CFB

0.20

I 0.18

I 4

I 8

I I I 12 16 20 = Pressure rotio, rp

I 24

I 28

I 32

Fig. 8 Effect of pressure ratio on steam cycle efficiency

0.t6

0.12

~

~

CFB

0.08

~ 0.0~,

I

0

I

I

I

I

I

1

I

I

8

12

16

20

24

28

32

----- Pressure ratio, rp Fig. 9 Effect of pressure ratio on gas cycle efficiency