Thermodynamic modeling and optimization of multi-pressure ... - NOPR

Journal of Scientific & Industrial Research SRINIVAS et al: OPTIMUM PRESSURE HEAT RECOVERY STEAM GENERATOR IN COMBINED CYCLE Vol. 67, October 2008, pp.827-834

827

Thermodynamic modeling and optimization of multi-pressure heat recovery steam generator in combined power cycle T Srinivas1*, A V S S K S Gupta2 and B V Reddy3 1 2 3

School of Mechanical and Building Sciences, VIT University, Vellore 632 014

Department of Mechanical Engineering, J N T U College of Engineering, Kukatpally, Hyderabad 500 080

Faculty of Engineering and Applied Sciences, University of Ontario Institute of Technology, Oshawa, ON, L1H 7K4, Canada Received 27 September 2007; revised 17 July 2008; accepted 21 July 2008

Optimum configuration for single pressure (SP), dual pressure (DP) and triple pressure (TP) heat recovery steam generator (HRSG) is presented to improve heat recovery and thereby exergy efficiency of combined cycle. Deaerator was added to enhance efficiency and remove dissolved gases in feed water. A new method was introduced to evaluate low pressure (LP) and intermediate pressure (IP) in HRSG from local flue gas temperature to get minimum possible temperature difference in heaters instead of a usual fixation of pressures. Optimum location for deaerator was found at 1, 3 and 5 bar respectively for SP, DP and TP in heat recovery at a high pressure (HP) of 200 bar. Results also showed optimum pressure for air compression and steam reheater by means of three categories of heat recovery. Key words: Combined cycle, Deaerator, Exergy analysis, Heat recovery, Multi pressure, Steam generator

Introduction Combined cycle (CC) power plants are gaining wider acceptance due to more and more availability of natural gas, higher overall thermal efficiencies, peaking and two-shifting capabilities, fast start-up capabilities and lesser cooling water requirements. Optimization of heat recovery steam generator (HRSG) is particularly interesting for combined plants design in order to maximize work obtained in vapor cycle. Pasha & Sanjeev1 presented a discussion about parameters that influence type of circulation and selection for HRSG. Ongiro et al2 developed a numerical method to predict performance of HRSG for design and operation constraints. Ganapathy et al3 described features of HRSG used in Cheng cycle system, where a large quantity of steam is injected into a gas turbine to increase electrical power output. Subrahmanyam 4 discussed factors affecting HRSG design for achieving highest CC efficiency with cheaper, economical and competitive designs. Noelle & Heyen5 designed oncethrough HRSG, which is ideally matched to very high temperature and pressure, well into supercritical range. *Author for correspondence E-mail: [email protected]

Ragland & Stenzel6 compared four plant designs using natural gas with a view of cost benefits achieved through HRSG optimization. Casarosa et al7 determined operating parameters means both of a thermodynamic and of a thermoeconomic analysis, minimizing a suitable objective function by analytical or numerical mathematical methods. Optimal design and operation of a HRSG is possible with minimization of entropy generation8. Reddy et al9 applied second law analysis for a waste HRSG, which consists of an economizer, an evaporator and a super heater. Multi pressure steam generation in HRSG of a combined power plant improves performance of plant10. Pelster et al11 compared results of a reference CC with dual and triple pressure HRSGs and also with and without steam reheating models. Bassily12,13 modeled a dual and triple pressure reheat CC with a preset in constraints on minimum temperature difference for pinch points, temperature difference for superheat approach, steam turbine inlet temperature and pressure, stack temperature, and dryness fraction at steam turbine outlet without a deaerator in steam bottoming cycle. This study presents optimization of single pressure (SP), dual pressure (DP) and triple pressure (TP) of

828

J SCI IND RES VOL 67 OCTOBER 2008

HRSG in CC with a deaerator. Variations in CC exergy efficiency are plotted with compressor pressure ratio, gas turbine inlet temperature, high pressure (HP) in HRSG, steam reheat pressure, deaerator pressure and pinch point (PP). Optimized pressures for combustion, HP steam, steam reheater and deaerator for each of HRSG configuration are presented.

Materials and Methods Thermodynamic Model of Combined Cycle with SP, DP and TP HRSG

Optimized configuration in HRSG improves steam generation rate and hence steam turbine output. Simple gas cycle gives higher efficiency at low compressor ratio compared to intercooled-reheat gas cycle, thereby in this work simple gas cycle is selected as a topping cycle for CC 14. Steam cycle with a reheater is adopted as a bottoming cycle in CC, which is modeled separately with SP, DP and TP level configurations in HRSG (Fig. 1). Temperature-heat transferred diagram for each pressure level is depicted in Fig. 2. In each configuration of HRSG, a deaerator is located to gain higher efficiency as well as to remove dissolved gases in feed water. Condensate preheater (CPH) is located in HRSG as the last heat transfer surface to improve heat recovery. In this work, steam turbine inlet temperatures, low pressure (LP) and intermediate pressure (IP) are not get fixed as in a regular manner. Steam temperature and pressures are determined with local flue gas temperature in heaters. Pinch point (PP)-minimum temperature difference between gas turbine exhaust leaving evaporator and saturation temperature of steam in evaporators and terminal temperature difference (TTD)-temperature difference between exhaust entry and super heated steam in super heaters are maintained constant. Zero approach point (temperature difference between saturation temperature of steam and incoming water temperature) is assumed in economizer.

Thermodynamic Analysis of Combined Cycle

Assumptions for analysis of CC are tabulated (Table 1). Energy efficiency of CC is determined based on lower heating value (50, 145 kJ/kg) of fuel. Net work output (%) of CC of standard chemical exergy of fuel (52, 275 kJ/kg) is expressed as exergy efficiency of CC 15. For an isentropic compression process, entropy change for air is zero, ∆s = 0.

s O2 , 31′ − sO2 ,30 + 3.76( s N

' 2 , 31

 P   P  − s N 2 ,30 ) − R log 31  + 3.76 log 31   = 0 P  P30     30 

…(1)

Temperature of air after isentropic compression is estimated from iteration of Eq. (1). Actual temperature of compressed air is determined from compressor isentropic efficiency. Combustion equation in gas turbine combustion chamber is CH4 + x (O2 + 3.76 N2) → CO2 + 2 H2O + (x - 2)O2 + 3.76x N2

...(2)

In Eq. (2), x is amount of air to be supplied. Air supply is determined by energy balance of Eq. (2) to get required combustion temperature. ∆s for isentropic expansion in gas turbine is ∆s = 0 = s32 – s33'

(3)

…

where, s32 = [sCO + 2 sH O + (x - 2) sO 2

2

2

+ 3.76 x sN2] 32 - R [log (P32/ mpcc) + 2 log(2 P32)/ mpcc)+ 3.76x log (3.76x P32/ mpcc) + (x - 2) log ((x - 2) P32)/ mpcc)]

...(4)

and s332 = [sCO + 2 sH O + (x - 2) sO + 3.76.xsN2]33' 2

2

2

- R [log (P33/ mpcc) + 2 log (2 P33 / mpcc) + 3.76x log (3.76x P33/ mpcc) + (x - 2) log ((x - 2) P33)/ mpcc)]

...(5)

where mpcc (kg mol) is total mass of chemical elements in products of combustion. Gas temperature after expansion in gas turbine is determined with iteration of entropy in Eq. (3). Actual temperature of expanded gas is obtained from gas turbine isentropic efficiency. IP and LP are evaluated from saturation temperatures. Steam flow rates and local exhaust temperature in heating devices are determined from heat balance equations. Saturation temperature of IP evaporator is TIP sat = TIP ex out – TTDIP – DSHIP

…(6)

829

SRINIVAS et al: OPTIMUM PRESSURE HEAT RECOVERY STEAM GENERATOR IN COMBINED CYCLE

31 32

Fuel

Atm. air 30

Fuel CC

32

31 30

CC

Atm. air G

G 33

GT

38

HRSG SH RH 34

33

RH

LP ECO CPH

34

35

36

22

21

37

12

39

38

9

13

HP LT ECO

LP EVAP

HP HT ECO

HPSH

EVAP ECO CPH 35 36 37

42

LP SH

HP EVAP

40

41

16

17 20

3 11

G

ST

19

5

LPST

1 HPST

3

6 Condenser

2

HP FP

G

2

8 6

4

12

Condenser

Deaerator

FP

7 Deaerator 10

18

4

1

11

14

8

FP

CEP LP FP

15

13

(a)

(b) Fuel

31

32

30

IP EVAP

HP HT ECO

HP SH

34

CC

G

GT

33

RH

Atm. air

HP EVAP

35

36

37

1

38 21 22

25

26

4

LP SH

HP LT ECO

IP SH

5

39

LP ECO

40

41

42

43

20

16

10

15

44

19

24

8

Exhaust

LP EVAP

IP ECO

9

3

G 2 6

11 Condenser

17

Deaerator

13

LP FP 14

18 19

IP FP

HP FP

23

(c)

Fig. 1—Schematic representation of combined cycle with: a) Single pressure HRSG; b) Dual pressure HRSG; and c) Triple pressure HRSG (HP, high pressure; IP, intermediate pressure; LP, low pressure; HT, high temperature; LT, low temperature; GT, gas turbine; ST, steam turbine; RH, reheater; SH, superheater; EVAP, evaporator; ECO, economizer; CPH, condensate preheater; FP, feed pump; HRSG, heat recovery steam generator; CEP, condensate extraction pump)

830


Fig. 2—Temperature-transferred heat diagram for: a) Single pressure; b) Dual pressure; and c) Triple pressure HRSG in combined cycle

Total net work output by CC, wnet, cc = wnet, gc + wnet, sc …(11)

Saturation temperature of LP evaporator is TLP sat = TLP ex out – TTDLP – DSHLP

…(7)

At HP evaporator, outlet temperature of flue gas = THP sat + PPHP

(8)

…

At IP evaporator, outlet temperature of flue gas = TIP sat + PPIP

To determine exergetic losses, irreversibilities associated in all components have to be estimated for exergy analysis. In this analysis, efficiency and losses are determined for 1 kg mol of fuel. Chemical exergy and physical exergy can be determined at each state as

…(9)

At LP evaporator, outlet temperature of flue gas = TLP sat + PPLP …(10) Work outputs and inputs in gas and steam cycles are related to unitary mass flow of fuel.

Chemical exergy, ech = ∑ k n k ε 0 k + RT0 Σ k n k ln .[ P .x k ] where xk is mole fraction of kth component

…(12)

831


Table 1—Assumptions made for thermodynamic evaluation of HRSG in combined cycles Atmospheric condition

25°C and 1.01325 bar

Gas cycle pressure ratio and maximum temperature

12 and 1200°C

Inlet pressure for HP steam turbine

200 bar

Terminal temperature difference (TTD) -temperature difference between flue gas and reheated/superheated steam

25

Condenser pressure

0.05 bar

Steam reheat pressure

50 % of HP pressure

Pinch Point (PP) - minimum temperature differences between flue gas and steam in evaporators of HRSG

25

Degree of superheat (DSH) in superheater

50

Isentropic efficiency of gas turbine

90 %.

Isentropic efficiencies of compressor and steam turbine

85 %

Pressure drop in combustion chamber

5 % of combustion chamber pressure

Pressure drop in HRSG, deaerator and condenser is neglected; Heat loss in HRSG, turbines, condenser, and deaerator is neglected

Exergetic loss in condenser, Physical exergy = eph = h - Σ k T0 s k Exergy, e = ech + eph

…(13) …(14)

Exergetic loss (irreversibility) in compressor, ic = e30 + wc - e31 (15) … Exergetic loss in gas turbine combustion chamber, igtcc = eCH4 + e31 - e32 …(16)

  T w , out i co = T 0  m co , st ( s out − s in ) + m w × 4 .18 × log    T w , in  

   

…(21) Exergetic loss of hot water from condenser,  Tw, out   iw ,ic =m w ×4.18 × ( Tw , out − Tw , in ) − T0 mw × 4.18 × log  T w , in   

…(22)

Exergetic loss in gas turbine, igt = e32 - e33 - wgt…(17)

Total (internal + external) exergetic loss in condenser, icot = ico+ iw …(23)

Exergetic loss in HRSG, iHRSG = (eHRSG in - eHRSG out) +Σ (ewater in - esteam out)

…(18)

Exergetic loss in exhaust from HRSG, iex = eex

…(19)

Exergetic loss in steam turbines, ist = T0 Σ m(sout - sin)

…(20)

Exergy loss in deaerator, ide = T0 [mde steam (safter mix – sbefore mix) + mfeed water (safter mix - sbefore mix)]

…(24)

Total exergetic loss of combined cycle,

832


Fig. 3-Exergy efficiency of combined cycle with different configurations (SP, DP and TP) of HRSGs versus: (a) Compressor pressure ratio; (b) Gas turbine inlet temperature; (c) HP pressure; (d) Steam reheater pressure; (e) Deaerator pressure; and (f) Pinch point

itot =

ic + igtcc + igt + iHRSG + iex + ist + icot + ide …(25)

Exergy efficiency of combined cycle,

 wnet cc η 2 , cc =   ε0  CH 4

  × 100  

…(26)

Results and Discussion CC with SP, DP and TP HRSG has been studied parametrically to find maximum obtainable exergy efficiency from CC. Deaerator pressure, at which steam is separated from steam turbine to heat feed water, is a key parameter to get optimum heat recovery from


833

Table 2—Comparison steam properties in three HRSGs and performance (rc = 12, Tg = 6130C, mg = 650 kg/s and mfuel = 13.1 kg/s) SP HRSG HP Steam 83.5 kg/s Pressure 200 bar Temperature 525°C

Net output: Gas cycle 236 MW Steam cycle 125 MW Combined cycle 361 MW Energy efficiency 55 %

DP HRSG HP Steam 83.5 kg/s Pressure 200 bar Temperature 525°C

TP HRSG HP Steam 83.5 kg/s Pressure 200 bar Temperature 525°C

LP Steam Pressure Temperature

IP Steam Pressure Temperature LP Steam Pressure Temperature Net output: Gas cycle Steam cycle Combined cycle Energy efficiency

16.7 kg/s 8.5 bar 230°C

Net output: Gas cycle 236 MW Steam cycle 140 MW Combined cycle 376 MW Energy efficiency 57.2 %

Fig. 4—Comparison of exergetic losses in combined cycle components with SP, DP and TP HRSG

exhaust. Effects of topping cycle parameters (compressor pressure ratio and gas turbine inlet temperature) and bottoming cycle parameters (HP pressure, steam reheater pressure, deaerator pressure and PP in HRSG) on exergy efficiency of CC are plotted to identify maximum efficiency. Optimum pressure ratio increases with increase in pressure level in HRSG (Fig. 3a). Gas turbine outlet temperature decreases with increase in pressure ratio. Therefore, efficiency of cycle decreases with further increase in pressure ratio from optimum value. Heat

17 kg/s 12 bar 239°C 15 kg/s 2 bar 165°C 236 MW 147 MW 383 MW 58.3 %

recovery improves from SP to TP configuration and hence efficiency of CC also rises with single to multi pressure effect. Optimum pressure ratio with SP, DP and TP HRSGs is 8, 10 and 12 respectively at gas turbine inlet temperature of 1200°C. At a fixed compressor ratio of 12, for SP, DP and TP configurations, heat recovery and exergy efficiency of CC increases with increase in temperature (Fig. 3b). But, at high temperature, there is a less increment in efficiency with multi pressure HRSG compared to at low temperature. Efficiency of cycle increases with increase in steam turbine inlet pressure for all configurations in HRSG (Fig. 3c). There is a significant increase in efficiency from DP to TP effect at high pressure (HP). At a fixed steam turbine inlet pressure, 200 bar, SP HRSG exergy efficiency of CC increases all time with increase in steam reheat pressure (Fig. 3d). A steam reheater is added to increase dryness fraction of steam at turbine exit and to avoid erosion of blades during wet expansion. For DP and TP HRSGs, exergy efficiency maximizes at optimum steam reheater pressure of 100 bar with HP pressure at 200 bar. Optimum deaerator pressure with SP, DP and TP levels in HRSG is around 1, 3 and 5 bar respectively with HP pressure at 200 bar and steam reheating at 100 bar (Fig. 3e). Optimized exergy efficiency at these conditions is 52.5, 54.5 and 55.5 % with SP, DP and TP HRSGs respectively. Exergy efficiency decreases with increase in PP (Fig. 3f). Irreversibility losses in heat transfer between exhaust and steam increases with increase in temperature difference. Heat recovery also

834


drops with increase in temperature difference (PP). Therefore, at low PP, higher efficiency for CC can be achieved. But, close temperature difference increases size of heating device and hence the cost. Their choice is the consequence of a compromise between thermodynamic efficiency and investment costs16. Thus, gain in efficiency from SP to DP is more compared to the gain from DP to TP level in HRSG. For SP, DP and TP heat recovery, compressor pressure ratio and gas turbine inlet temperature are fixed. Therefore, topping cycle components exhibit no change on exergetic losses, which are compressor, gas combustion chamber and gas turbine. Steam generation rate increases from SP to TP in HRSG, thereby exergetic loss associated with heat transfer increases (Fig. 4). Exhaust temperature and hence exergetic loss in exhaust decreases from SP to TP levels in HRSG. Exergetic loss in expansion of steam in turbine rises with pressure levels due to excessive steam flow rate. This also increases condenser exergetic loss with increase in pressure levels. More amount of steam is required for feed water heater in case of SP HRSG whereas in DP and TP effect, amount of bled steam and hence exergetic loss in deaerator decreases. On overall basis, total exergetic loss decreases from SP to TP HRSG. Keeping exhaust gas flow rate fixed (650 kg/s), results (Table 2) for SP, DP and TP heat recovery agreed with literature values7. Conclusions Optimum pressure ratio for compressor with SP, DP and TP effects in heat recovery are 8, 10 and 12 respectively at 1200°C of gas turbine inlet temperature. Optimum deaerator pressure is obtained at 1, 3 and 5 bar for SP, DP and TP levels respectively at steam turbine inlet pressure of 200 bar. Similarly, at 200 bar of HP pressure for DP and TP, steam reheater demands 100 bar to maximize exergy efficiency for CC. Parametric analysis exhibits that gain in efficiency form single pressure heat recovery to DP and TP recovery increasing with diminishing rate.

References 1

2

3

4

5

6

7

8

9

10

11

12

13

14

15 16

Pasha A & Sanjeev J, Combined cycle heat recovery steam generators optimum capabilities and selection criteria, Heat Recovery systems & CHP, 15 (1995) 147-154. Ongiro A, Ugursal VI, Taweel A M & Walker J D, Modeling of heat recovery steam generator performance, Appl Thermal Engg, 17 (1997) 427-446. Ganapathy V, Heil B & Rentz J, Heat recovery steam generator for Cheng cycle application, in ASME Ind Power Conf – PWR, San Diego, 4 (1988) 61-65. Subrahmanyam N V R S S, Rajaram S & Kamalanathan N, HRSGs for combined cycle power plants, Heat Recovery Systems & CHP, 15 (1995) 155-161. Noelle M & Heyen G, Mathematical modeling and design of an advanced once-through heat recovery steam generator, Compu Chem Engg, 28 (2004) 651-660. Ragland A & Stenzel W, Combined cycle heat recovery optimization, in ASME Proc 2000 Int Joint Power Generation Conf, IJPGC2000-15031 (Miami Beach, Florida) 23-26 July 2000, 1-6. Casarosa F, Donatini & Franco A, Thermoeconomic optimization of heat recovery steam generators operating parameters for combined plants, Energy, 29 (2004) 389-414. Nag P K & De S, Design and operation of a heat recovery steam generator with minimum irreversibility, Appl Thermal Engg, 17 (1997) 385-391. Reddy B V, Ramkiran G, Kumar A K & Nag P K, Second law analysis of a waste heat recovery steam generator, Int J Heat Mass Transfer, 45 (2002) 1807-1814. De S & Biswal S K, Performance improvement of a coal gasification and combined cogeneration plant by multi-pressure steam generation, Appl Thermal Engg, 24 (2004) 449-456. Pelster S, Favrat D & Spakovsky M R, Thermoeconomic and environomic modeling and optimization of the synthesis, design, and operation of combined cycles with advanced options, ASME J Gas Turbines Power, 123 (2001) 717-726. Bassily A M, Modeling, numerical optimization, and irreversibility reduction of a dual-pressure reheat combined cycle, Appl Energy, 81 (2005) 127-151. Bassily A M, Modeling, numerical optimization, and irreversibility reduction of a triple-pressure reheat combined cycle, Energy, 32 (2007) 778-794. Srinivas T, Gupta A V S S K S & Reddy B V, Performance simulation of 210 MW natural gas fired combined cycle power plant, J Energy Heat Mass Transfer, 29 (2007) 61-82. Kotas T J, The Exergy Method of Thermal Plant Analysis (Krieger publishing company, Malabar, F L) 1995, 236-269. Franco A. & Russo A, Combined cycle plant efficiency increase based on the optimization of the heat recovery steam generator operating parameters, Int J Thermal Sci, 41 (2002) 843-859.