Front. Energy DOI 10.1007/s11708-015-0371-9

RESEARCH ARTICLE

Nikhil DEV, Rajesh ATTRI

Performance analysis of combined cycle power plant

© Higher Education Press and Springer-Verlag Berlin Heidelberg 2015

Abstract Combined cycle power plants (CCPPs) are in operation with diverse thermodynamic cycle configurations. Assortment of thermodynamic cycle for scrupulous locality is dependent on the type of fuel available and different utilities obtained from the plant. In the present paper, seven of the practically applicable configurations of CCPP are taken into consideration. Exergetic and energetic analysis of each component of the seven configurations is conducted with the help of computer programming tool, i.e., engineering equation solver (EES) at different pressure ratios. For Case 7, the effects of pressure ratio, turbine inlet temperature and ambient relative humidity on the first and second law is studied. The thermodynamics analysis indicates that the exergy destruction in various components of the combined cycle is significantly affected by the overall pressure ratio, turbine inlet temperature and pressure loss in air filter and less affected by the ambient relative humidity. Keywords first-law, second-law, exergy destruction, components

1

Introduction

India is swiftly developing its economy, with a need for independent and reliable supply of electricity and to be a power sufficient country is one of its prime concerns. The present installed capacity of electricity in India is 249488 MW, out of which, 172986 MW (July 31, 2014, CEA India) is produced by thermal power plants. The per capita consumption of electric power in India was 883.6 kWh as estimated on March 31, 2012 and 917.2 kWh on March 31, 2013, while the per capita consumption of power in developed countries like USA was 13,361 kWh in 2010.

Received December 11, 2014; accepted March 4, 2015

✉

Nikhil DEV ( ), Rajesh ATTRI Department of Mechanical Engineering, YMCA University of Science and Technology, Faridabad 121006, India E-mail: [email protected]

Therefore, installation of many combined cycle power plants is currently in progress to meet electricity demand in the future. Thermal power plants account for 69.34% of the installed electricity generation capacity. Out of the technologies for thermal power generation, gas turbine (GT) technology is a critical one for the future, providing a clean, efficient and cost effective way of generating power from distributed generation and co-production schemes for large utility-size combined cycle plant. There are diverse GT systems such as stand-alone, combined with a steam turbine (ST) (combined cycle) or, less commonly at present, with a fuel cell. The main markets are in the power generation industry, process industry, mobile applications (land and military marine), and gas and oil transmission pipeline pumping stations [1–12]. The main technical obstacles to the application of GT technology are that the stand-alone GT has a lower efficiency in its basic configuration than an equal power output reciprocating engine. Moreover, the efficiency decreases at partial load and burning of low heating value fuels may not be feasible, depending on the type of the turbine. The main non-technical obstacles to the application are that the maintenance entails more skilled personnel than does the reciprocating engine and that small GT are too expensive as compared to engines [11]. Natural gas use is expected to grow. However, this will not be accessible in all geographical areas. Therefore, in the industrial market, the fuels mix is expected to become more diverse including naphtha, kerosene, gas condensates, natural gas liquids, alcohols, refinery residues, biomass, etc. This fuel flexibility will of course need to be matched to a low emissions capability [9,10]. All major GT manufacturing companies, including General Electric, ABB-Alstom, Westinghouse Siemens and Mitsubishi offer last generation combined cycle plants of power range 680 MW, with very high efficiency claimed up to 62%. This value is very high indeed, in comparison with the Carnot factor value of about 0.8 for a turbine inlet temperature (TIT) of 1523 K, typical for the F series GT. Because of this higher efficiency, the GT and combined cycle power plants are becoming more and more attractive

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with regard to reduced fuel consumption and less emissions. India has been spending a great amount of money importing fuels, which could be used in the development and growth of the country [12]. To reduce the shortage of power, efforts are required to improve the performance of prevailing plants. For this purpose, it is vital to evaluate the thermodynamic performance of each component. It is reported in the literature that first law analysis alone is not sufficient enough to evaluate its performance. Exergetic analysis along with energetic analysis gives the true picture of thermal inefficiency of a system [8]. Exergy analysis is performed to identify the sources of work potential loss. In the design and analysis of thermal systems, this method is used to explore the proper parameters to minimize work potential. The analysis may be conducted either of the system or of the component depending on the objective of the optimization. Optimization of the component may not necessarily lead to the optimization of the overall system. The loss of work potential may either be caused by ‘exergy loss’ or by ‘exergy destruction’. Exergy loss may be defined as the loss of work potential due to the exergy transfer to the environment, i.e., unused exergy. On the other hand, exergy destruction is the loss of work potential due to the irreversibilities within a system. It is alternately named as availability destruction, irreversibility, lost work etc. in the literature. Thermal systems are typically supplied with the exergy inputs associated directly or indirectly with fossil fuels or other energy sources. Accordingly, the destruction and losses of exergy signify the waste of these energy resources. The method of exergy analysis targets at the quantitative evaluation of the exergy destruction and losses associated with a system. Exergy destruction in most of the thermal systems stem from one or more of the three principal causes, i.e., combustion/chemical reaction, heat transfer and friction. Combustion is intrinsically a very significant cause of exergy destruction. For typical atmospheric combustion systems, about 1/3rd of the fuel energy becomes unavailable and is discharged into the environment as heat. Most of this irreversibility is associated with the internal heat transfer within the combustor between the products and reactants. Such heat transfer becomes inevitable in both premixed and diffusion flames, where highly energetic product molecules are free to exchange energy with unreacted fuel and air molecules. However, the inefficiency of combustion may be reduced by preheating the combustion air and reducing the air fuel ratio. Exergy destruction associated with heat transfer decreases as the temperature difference between the streams is reduced. However, this may increase the size leading to greater friction loss. The present paper is concentrated on determination of the exergy destruction of each component of GT and combined cycle power plant for different configurations

and operating conditions. For the present analysis, the following seven homogeneous configurations are taken into consideration: Case 1: Simple GT; Case 2: GT + regenerator; Case 3: GT + regenerator + intercooler ; Case 4: GT + regenerator + intercooler + reheater (REH); Case 5: GT + regenerator + intercooler + REH + ST; Case 6: GT + regenerator + intercooler + REH + ST + 1 feed heater; Case 7: GT + regenerator + intercooler + REH + ST + 2 feed heater.

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System description

Figure 1 shows the schematic diagram of a combined cycle power plant with two feedwater heaters, an air filter and an inlet air cooling. Ambient air enters the air filter to protect the compressor blades and other components of the plant from the dust particle. After pressure drop at constant pressure in the air filter, the air reaches state 2. Ambient air is cooled and humidified in the air humidifier (AH1) before it counters pass the inlet air to the lower-pressure compressor (LPC) in the air cooler (AC), cooling the inlet air. The air at state 3 is cooled to a temperature close to the wet bulb temperature at state 2. The compressed air from the LPC at state 4 is cooled to state 5 in an indirect inter cooler using the ambient humidified air that is humidified in the air humidifier (AH2) and has a lower temperature than the ambient temperature. This air enters the high-pressure compressor (HPC) and is compressed from state 5 to state 6. The outlet air of HPC at state 6 enters the regenerative heat exchanger (HE) where it exchanges heat with power turbine (PT) exhaust at state 11, gets heated a statet 7 and supplied to the combustion chamber (CC) (where fuel is burned), producing hot gases at state 8. The hot gas is then expanded to state 9 in the high pressure turbine (HPT) to a lower pressure and temperature before it is recombusted in the REH, after which the reheated gas expands, through a PT to drive a load. Part of the heat of hot exhaust gas is used in the HE, and part of the heat is utilized in the waste heat recovery boiler (WHRB) to generate steam. The steam produced in WHRB from liquid feed by pump (P3) by exchanging heat enters the ST at state 13 and expands to condenser pressure at state 16 to drive a load. In the condenser, the ST exhaust is condensed to saturated state 17 and fed by the pump (P1) in the feedwater heater (FWH2) at state 18 for direct mixing with the steam which bled from the ST at state 15 to FWH2. The saturated liquid at state 19 is pumped by P2 to feed the water heater (FWH1) at state 20 for direct mixing with the steam bled from the ST at state 14 to FWH1. The saturated liquid at state 21 is fed by the pump (P3) to WHRB at state 22 to produced heat.

Nikhil DEV et al. Performance analysis of combined cycle power plant

For the TIT, the inlet temperature of the exhaust gas to the regenerator is limited by the HE materials. The use of stainless steel in recuperators has an upper temperature capability of approximately 1089 K. Nickel-based superalloys make the operation possible at 1123 K. More advanced oxide dispersion alloys have the potential to withstand temperatures of around 1223 K [1]. A ceramic HE made from silicon carbide composites might operate at even higher temperatures. A temperature of up to 1348 K was achieved in the experiments [2] and a temperature of 1643 K and even higher were also reported [6]. Improvement of GT efficiency and power output can be accomplished in many ways. One way is to increase the TIT and pressure ratio, however, this requires advanced materials and larger cooling flows for the GT hot parts. GT for stationary power generation have benefited from research and development on aircraft GTs, which have also been converted to power generation applications (Case 1). Another way to increase power output is compressor intercooling, which also raises the efficiency for cooled aero derivative gas turbines. Intercooling reduces the compression work and the cooling air temperature, thus enabling higher firing temperatures or lower cooling flow rates. Moreover, the GT efficiency can be augmented by recovering the exhaust gas energy, for example by using a recuperator that preheats the compressed air prior to the combustor, thus decreasing the fuel flow rate (Case 2). GT recuperators are relatively scarce and have mostly been used for propulsive GTs, for which low specific fuel consumption (kg/kWh) is important [3,4]. Intercooling combined with recuperation increases the efficiency and power output more than either

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method applied alone (Case 3). One example of an intercooled recuperated GT is the WR-21, which is used in the UK Royal Navy’s next generation warships [5]. The increase in efficiency results from the increase in specific power output, being the difference between output and input power. This, in turn, can be explained by separating the effect of compressors and turbines. The work input needed for compression is lower due to the intercooler. The difference in work input compared to the simple cycle increases with the pressure ratio. Obviously, the work output of a turbine increases with the pressure ratio. Compared to the simple cycle, the output increase augments because less cooling air is needed in the intercooled cycle. This explains the increase in the specific power as compared to the simple cycle and the higher increase in efficiency with higher pressure ratios. The most important parameter that affects the performance of a GT is the intake air temperature. The air mass flow rate entering a GT as a function of an ambient temperature can be simply computed based on the actual data. A higher air mass flow rate also increases fuel consumption and exhaust gas. To perform the comparative study, the configurations of other six cases are depicted in Figs. 2 to 7. A very interesting opportunity to increase the efficiency of the combined cycle plant is represented by exploring the concept of gas to gas recuperation and post-combustion, with the thermodynamic scheme shown in Fig. 5 (Case 4). The post-combustion concept has been already applied in the ABB GT24/26 series turbine but with a very high pressure ratio (30:1). In the open or direct-contact feedwater heater, the

Fig. 1 Schematic diagram of the combined cycle power plant (CCPP) with two feedwater heaters (Case 7)

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extraction steam is mixed directly with the incoming subcooled feedwater to produce saturated water at the extraction steam pressure. In this paper, the combined cycle power plant without steam extraction (Case 6) and with single steam extraction (Case 7) is studied for exergy destruction in different components.

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Mathematical model

Exergy is destroyed in all natural processes. This is known

Fig. 2 Schematic diagram of simple GT cycle (Case 1)

as the law of degradation of energy. Thermodynamic irreversibilities are the causes of exergy destruction. All natural processes are irreversible. The causes of irreversibility lie in the basic requirements for the natural processes to occur, which are the lack of thermodynamic equilibrium and the dissipative effects. The thermodynamic irreversibility is characterized by the entropy generation rate. Exergy analysis is the combination of the first and second laws of thermodynamics to evaluate the efficiencies of processes and devices. If the system operates in a steady-state, steady flow condition and all the nonreacting

Fig. 3

Schematic diagram of GT cycle with regenerator (Case 2)

Fig. 4 Schematic diagram of GT cycle with regenerator and intercooler (Case 3)

Fig. 5

Schematic diagram of GT cycle with regenerator, intercooler and REH (Case 4)

Nikhil DEV et al. Performance analysis of combined cycle power plant

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Fig. 6 Schematic diagram of CCPP (Case 5)

Fig. 7

Schematic diagram of CCPP with one feedwater heater (Case 6)

gases are arbitrarily assigned as zero thermo-mechanical enthalpy, entropy, and exergy at the condition of ambient pressure and temperature regardless of their chemical composition. Then the entropy of mixing different gaseous components can be neglected, and the general exergybalance equation is given by

:

EW ¼

n X

:

EQ i þ

i¼1

X

:

me –

X

:

:

me – ED :

(1)

out

in

For single stream flow, :

:

:

:

:

EW ¼ EQ þ m ein – m eout – m eD :

(2)

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The mathematical equations for different components of the CCPP are tabulated in Table 1. The ratio of all the useful energy extracted from the system (electricity) to the energy of fuel input is known as the first-law efficiency and defined by ηI ¼

Wel,CCPP , Hf

(3)

where Wel,CCPP is the electrical power output and is given by Wel,CCPP ¼ Wel,GT þ Wel,ST ,

(4)

Wel,GT ¼ ηgen Wnet,GT ,

(5)

Wel,ST ¼ ηgen Wnet,ST ,

(6)

Wnet,GT ¼ WHPT þ WPT – WLPC :

(7)

The word exergy is derived from the Greek ex (out or outer) and ergon (force or work). The concept was first noticed in 1824 by Carnot in the relation of heat and work. Exergy analysis combines the first and second laws of thermodynamics, and is a powerful tool for analyzing both the quantity and the quality of energy utilization. Exergy is defined as the maximum work obtainable while the system communicates with environment reversibly [7]. Since exergy is more valuable than energy according to the second law of thermodynamics. It is useful to consider both output and input in terms of exergy [8]. The amount of exergy supplied in the product to the amount of exergy associated with the fuel is a more accurate measure of the thermodynamic performance of a system which is defined as ηII ¼

Wel,CCPP , E_ f

(8)

where E_ f ¼ efCC mfCC þ efREH mfREH :

The program coding for this paper is written in Engineering Equation Solver (version 7.929). The flowchart for the evaluation of CCPP efficiency has been epitomized in Fig. 8.

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Assumptions

In this paper, the GT capacity is 30 MW. The ambient pressure and temperature are, respectively, 1 bar and 298 K. The fuel is methane gas, which has a lower heating value of 42000 kJ/kg. The remaining assumptions for the other components are as represented in Table 2.

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Results and discussion

Energy is conserved, while exergy is consumed due to irreversibilities. Exergy indicates the quality of energy, and in any real process, it need not be conserved, but it is destroyed or lost. Here, an application of the energy and exergy modeling technique discussed in the previous section is presented for the energy and exergy use in the electricity generation sector. The method of exergy analysis aims at the quantitative evaluation of the exergy destruction and losses associated within system. An exergy balance, by definition, only exists in reversible processes. In real processes, exergy is never in balance, because the total exergy input always exceeds the total exergy output. By calculating the sum of exergy loss and destruction, possible process improvements are visualized. In general, when the exergy loss is high, this part should be considered for improvement first. However, this strategy is not always appropriate. The reason is that each part of the system

Table 1 Exergy destruction rate equations for plant components Components

Exergy destruction rate eD,AC ¼ einAC – eoutAC þ W_ AC

Air compressor CC

eD,CC ¼ einCC þ ef – eoutCC

GT

eD,GT ¼ ðeinGT – eoutGT Þ – W_ GT

HE WHRB Air filter Air humidifier ST Feedwater heater first

eD,HE ¼ mair ðeaiHE – eaoHE Þ þ mgas ðegiHE – egoHE Þ eD,WHRB ¼

mgas ðeginWHRB – egoutWHRB Þ – mw ðewoutWHRB – ewinWHRB Þ eD,AF ¼ ma ðeaiAF – eaoAF Þ eD,AH ¼ maiAH eaiAH þ mw ew – maoAH eaoAH eD,ST ¼ ms ðewinST – ewoutST Þ – WST eD,FWH1 ¼ ms1 e14 þ ðms – ms1 Þe20 – ms e21

Feedwater heater second

eD,FWH2 ¼ ms2 e15 þ ðms – ms1 – ms2 Þe18 – ðms – ms1 Þe19

Condenser

eD,COND ¼ ðms – ms1 Þðewin – ewout Þ þ mcw ðecwin – ecwout Þ

Pump

(9)

eD,P ¼ WP þ ms ðewin – ewout Þ

Nikhil DEV et al. Performance analysis of combined cycle power plant

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Fig. 8 Flowchart for CCPP efficiency evaluation

depends on the other parts, so that an improvement in one part may cause increased losses in other parts. Some performance parameters are defined to estimate the exergetic performance. Figure 9 shows the variation of first-law efficiency of CCPP for Case 7 with a change in compressor pressure ratio rp for different values of TIT and constant ambient relative humidity φ = 60%. As the pressure ratio increases, the compressor work increases, raising the temperature at the compressor outlet. Increase the pressure ratio also increases the turbine work. As the pressure ratio increases, the air temperature at the inlet of the CC decreases due to the decrease in the temperature of turbine exhaust entering the regenerative HE. The ratio of network output of the CCPP to the heat added represent the first-law efficiency of the cycle increase to its maximum value at optimum rp for a particular TIT. Further increase in rp reduces the first-law efficiency because, at a much higher pressure ratio, the heat added to the cycle increases. Figure 10 displays the variation of second-law efficiency of CCPP for Case 7 with a change in rp for different values of TIT and φ = 60%, which is a more accurate measure of

the thermodynamic performance. Since the quality of fuel (i.e., exergy associated with the heat addition) is more than the heating value or energy of fuel because the exergy of fuel would increase while bringing it from the ambient pressure to combustion pressure at ambient temperature. Hence, the exergy associated with the heat addition will be equal to the exergy associated with the heating value of the fuel plus exergy increase, i.e., mechanical exergy due to the increase of pressure of the fuel from the ambient to combustion state. Therefore, the second-law efficiency of cycle is slightly lower than the first-law efficiency. The optimum pressure ratio is found to be increased with the increase in TIT (Fig. 11) It is found that the exergy destruction of the CC is higher than that of other plant components, due to the high irreversibility in the former. Figure 12 shows the variation of the exergy destruction in the CC for all cases with respect to the change in pressure ratio at a fixed value of TIT = 1400 K and φ = 60%. It is found from the analysis that the exergy destruction in the combustion process dominates. It represents over 60% of the total exergy destruction in the overall system. As the pressure ratio

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Table 2 Assumptions for different components of the cycles Component

Quantified assumption

Pressure losses in air filter at 100% air flow/mbar

3.5

Relative humidity at air humidifiers outlet/%

100

Pressure drop for air in the air intercooler/%

1

The pressure drop for air in the AC/%

1

Pressure drop for gas in the regenerative HE/%

2

Pressure drop in the CC and REH/%

4

Pressure drop in the waste heat recovery boiler/%

4

Effectiveness for the AC/%

85

Effectiveness of the air intercooler/%

90

Effectiveness of the regenerative HE/%

55

Compressor isentropic efficiency/%

87

GT isentropic efficiency/%

89

Efficiency of the CC and REH/%

95

Generator efficiency/%

97

Steam pressure at the ST inlet/bar

25

Steam temperature at the ST inlet/K

567

ST exhaust pressure/bar Temperature rise of cooling water in condenser/K

Fig. 10 Effect of variation of pressure ratio at various TIT on second-law efficiency

0.09 283.98

Cooling water inlet temperature in condenser/K

298

Cooling water outlet temperature from condenser/K

309

Stack temperature/K

413

Pump isentropic efficiency/%

85

Turbine isentropic efficiency/%

85

Fig. 11 Effect of TIT on optimum pressure ratio

Fig. 12 Effect of variation of pressure ratio on exergy destruction in CC Fig. 9 Effect of variation of pressure ratio at various TIT on firstlaw efficiency

increases, the exergy destruction in the CC decreases significantly except for Case 2 and Case 3. This is due to the increase in the pressure ratio which implies a smaller difference of exergy between the combustion products and compressed air but it drops with the exergy carried by fuel. Also, as the pressure ratio increases, the exergy loss in combustion or the reaction decreases due to the decrease in the mass of fuel. As the pressure ratio increases, the mass flow rate of air also decreases.

For Case 2, rp is taken from 6 to 20, because at the pressure ratio above 20, the function of the regenerator fails as the temperature of the air at the compressor outlet becomes greater than the turbine exhaust temperature. For Cases 2 and 3, as the pressure ratio increases, the air rate decreases to a minimum value first and then increases [6]. As is known that the air rate is proportional to the mass flow rate of air, the mass of fuel in the CC also follows the same trend of decrease and then increases as the mass flow rate of air increases. Therefore, the exergy destruction in the CC decreases to a minimum value first and then increases. Figure 13 shows the variation of exergy

Nikhil DEV et al. Performance analysis of combined cycle power plant

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destruction in the CC for all cases at a fixed value of TIT = 1400 K, rp = 14 and φ = 60%. From Fig. 13, it is observed clearly that as the complicacy of the plant increases, the exergy destruction in the CC decrease. This means that the experiment is moving to the reversible process. For Cases 5 to 7, the exergy destruction in the CC is the same, because on adding any component on the ST, the exergy destruction in the CC is not affected. The exergy destruction in the CC for Cases 5, to 7 is greater than that for Case 4 due to the pressure loss in the air filter, WHRB and functioning of other components.

Fig. 14 Effect of variation of pressure ratio on exergy destruction in regenerative HE

Fig. 13 Exergy destruction in CC at TIT = 1400 K, rp = 14 and φ = 60%

Figure 14 shows the variation of exergy destruction in the regenerative HE for Cases 2 to 7 with respect to the change in pressure ratio at a fixed value of TIT = 1400 K and φ = 60%. The exergy destruction in the regenerative heat exchanger decreases as the pressure ratio increases. This is caused by the higher pressure ratio which results in a lower power turbine exit temperature. Figure 15 shows the variation of exergy destruction in the regenerative HE for Cases 2 to 7 at a fixed value of TIT = 1400 K, rp = 14 and φ = 60%. It is observed that the exergy destruction in the HE for Case 4 is greater than that for Case 3. The reason for this is that at the same pressure ratio and TIT, the power turbine exhaust temperature for Case 4 is greater than the turbine exhaust temperature for Case 3. For Cases 5 to 7, the exergy destruction in the HE are the same, because on adding any component on the steam turbine, the exergy destruction in the HE is not affected. The exergy destruction in the HE for Cases 5 to 7 is greater than that for Case 4 due to the pressure loss in the air filter, WHRB and functioning of other components. Figure 16 shows the variation of the exergy destruction in compressor for all cases with respect to the change in pressure ratio at a fixed value of TIT = 1400 K and φ = 60%. As the pressure ratio increases, the work of the compressor increases, so does the exergy destruction in the compressor for Cases 1 to 3. The air rate for Case 2 is greater than that for Case 1 [6]; therefore, the exergy destruction in the compressor for Case 2 is greater than that for Case 1.

Fig. 15 Exergy destruction in regenerative HE at TIT = 1400 K, rp = 14 and φ = 60%

Figure 17 shows the variation of exergy destruction in the compressor for all cases at a fixed value of TIT = 1400 K, rp = 14 and φ = 60%. It is seen from Fig. 17 that the exergy destruction in the high pressure compressor (HPC) is greater than that in the lower pressure compressor (LPC) for Cases 3 to 7, because the HPC consumes more work than does the LPC. The exergy destruction in HPC for Cases 5 to 7 and exergy destruction in LPC for the same cases (Cases 5 to 7) are the same, because on adding any component in ST, the exergy destruction in both compressors are not affected. The exergy destruction in LPC and HPC for Cases 5 to 7 is greater than that for Case 4, due to the effect of the pressure loss in the air filter, WHRB and functioning of other components. As the complicacy of the plant increases, this means that the experiment is moving to the reversible process as shown in Fig. 17. Figure 18 shows the variation of the exergy destruction in the intercooler for Cases 3 to 7 with respect to the change in pressure ratio at a fixed value of TIT = 1400 K and φ = 60%. As the pressure ratio increases, the exergy destruction in the intercooler increases as the temperature of the air at the inlet of the intercooler increases. Figure 19 shows the variation of exergy destruction in the intercooler for Cases 3 to 7 at a fixed value of TIT = 1400 K, rp = 14

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Fig. 16

Effect of variation of pressure ratio on exergy destruction in compressor

Fig. 17 Exergy destruction in compressor at TIT = 1400 K, rp = 14 and φ = 60%

Fig. 18 Effect of variation of pressure ratio on exergy destruction in intercooler

and φ = 60%. It is noticed from Fig. 19 that the exergy destruction in the intercooler for Case 4 is less than that for Case 3 as the mass flow rate of the air for Case 4 is less than that for Case 3. The exergy destruction in the intercooler

for Cases 5 to 7 are the same and greater than that for Case 4 due to the pressure loss in the air filter, WHRB and functioning of other components. Figure 20 shows the variation of exergy destruction in the turbine for all cases with respect to the change in pressure ratio at a fixed value of TIT = 1400 K and φ = 60%. As the pressure ratio increases, the exergy destruction in turbine for all cases increases, because the entropy at turbine inlet decreases. Figure 21 shows the variation of exergy destruction in the turbine for all cases at a fixed value of TIT = 1400 K, rp = 14 and φ = 60%. It is observed from Fig. 21 that the exergy destruction in the HPT is slightly more than that of the PT for Cases 3 to 7, because the difference of entropy at the outlet and inlet of the HPT is slightly more than that of the PT. The exergy destruction in the turbine in Case 2 is more than that in Case 1 as the air rate for Case 2 is more than that for Case 1. The exergy destruction in the HPT and PT for Cases 5 to 7 is slightly more than that for Case 4 due to the pressure drop in air filter, WHRB and functioning of other components.

Nikhil DEV et al. Performance analysis of combined cycle power plant

Fig. 19 Exergy destruction in intercooler at TIT = 1400 K, rp = 14 and φ = 60%

Figure 22 shows the variation of the exergy destruction in the REH for Cases 4 to 7 with respect to the change in pressure ratio at a fixed value of TIT = 1400 K and φ = 60%. It is found from the analysis that after the CC the REH has maximum exergy destruction in the plant. As the pressure ratio increases, the mass of fuel in the REH

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increases to compensate for the reduction in energy due to the decrease in the Tav. Due to the decrease in the Tav and increase in the mass of fuel in the REH and ratio A, the exergy destruction in the REH increases. Figure 23 shows the variation of exergy destruction in the REH for Case 4 to Case 7 at a fixed value of TIT = 1400 K, rp = 14 and φ = 60%. From this figure it is clear that the exergy destruction in the REH for the Case 5, Case 6 and Case 7 are same and is greater than that of 4 due to the pressure loss in air filter, WHRB and function of other components. Figure 24 shows the variation of exergy destruction in the air filter for Cases 5 to 7 with respect to the change in pressure ratio at a fixed value of TIT = 1400 K and φ = 60%. As the pressure ratio increases, the mass flow rate of the air decreases, so does the exergy destruction in the air filter. Figure 25 shows the variation of exergy destruction in the air filter for Cases 5 to 7 at a fixed value of TIT = 1400 K, rp = 14 and φ = 60%. It is seen that the exergy destruction in the air filter for Cases 5 to 7 are the same. Figure 26 shows the variation of exergy destruction in

Fig. 20

Effect of variation of pressure ratio on exergy destruction in turbine

Fig. 21

Exergy destruction in turbine at TIT = 1400 K, rp = 14 and φ = 60%

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Fig. 25 Exergy destruction in air filter at TIT = 1400 K, rp = 14 and φ = 60% Fig. 22 Effect of variation of pressure ratio on exergy destruction in REH

Fig. 23 Exergy destruction in REH at TIT = 1400 K, rp = 14 and φ = 60%

Fig. 26 Effect of variation of pressure ratio on exergy destruction in WHRB

Fig. 27 Exergy destruction in WHRB at TIT = 1400 K, rp = 14 and φ = 60% Fig. 24 Effect of variation of pressure ratio on exergy destruction in air filter

the WHRB for Cases 5 to 7 with respect to the change in pressure ratio at a fixed value of TIT = 1400 K and φ = 60%. As the pressure ratio increases, the PT exhaust temperature decreases, so does the exergy destruction in WHRB. Figure 27 shows the variation of exergy destruction in the WHRB for Cases 5 to 7 at a fixed

value of TIT = 1400 K, rp = 14 and φ = 60%. It is noticed that on moving from Case 5 to Case 7 the exergy destruction in the WHRB decreases. The reason for this is that the mass, temperature and exergy of water inlet to WHRB decrease. Figure 28 shows the variation of exergy destruction in the ST for Cases 5 to 7 with respect to the change in pressure ratio at a fixed value of TIT = 1400 K and φ = 60%. As the pressure ratio increases, the mass of the steam

Nikhil DEV et al. Performance analysis of combined cycle power plant

Fig. 28 Effect of variation of pressure ratio on exergy destruction in ST

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Fig. 30 Effect of variation of pressure ratio on exergy destruction in condenser

produced and the work of the ST decrease, so does the exergy destruction in the ST. Figure 29 shows the variation of exergy destruction in the ST for Cases 5 to 7 at a fixed value of TIT = 1400 K, rp = 14 and φ = 60%. It is observed that on moving from Case 5 to Case 7, the exergy destruction in the ST increases. This is because the mass of the steam and the work of the ST increase.

Fig. 31 Exergy destruction in condenser at TIT = 1400 K, rp = 14 and φ = 60%

Fig. 29 = 60%

Exergy destruction in ST at TIT = 1400 K, rp = 14 and φ

Figure 30 shows the variation of exergy destruction in the condenser for Cases 5 to 7 with respect to the change in pressure ratio at a fixed value of TIT = 1400 K and φ = 60%. As the pressure ratio increases, the exergy destruction in the condenser decreases because the mass of the steam inlet to the condenser and the mass of cooling water decrease. Figure 31 shows the variation of exergy destruction in the condenser for Cases 5 to 7 at a fixed value of TIT = 1400 K, rp = 14 and φ = 60%. It is noticed that on moving from Case 5 to Case 7, the exergy destruction in the condenser decreases due to the same reason as that in Fig. 30. Figure 32 shows the variation of exergy destruction in the feedwater heater for Cases 6 and 7 with respect to the change in pressure ratio at a fixed value of TIT = 1400 K

and φ = 60%. As the pressure ratio increases, the exergy destruction in the feedwater heater decreases due to the amount of mass of the steam entering and leaving the feedwater heater. Figure 33 shows the variation of exergy destruction in the feedwater heater for Cases 6 and 7 at a fixed value of TIT = 1400 K, rp = 14 and φ = 60%. It is seen that the summation of exergy destruction of both feedwater heaters in Case 7 is less than that in feed the water heater for Case 6. Figure 34 shows the variation of exergy destruction in the pump for Cases 5 to 7 with respect to the change in pressure ratio at a fixed value of TIT = 1400 K and φ = 60%. As the pressure ratio increases, the exergy destruction in the pump decreases as the mass of the steam entering the pump decreases. Figure 35 shows the variation of exergy destruction in the pump for Cases 5 to 7 at a fixed value of TIT = 1400 K, rp = 14 and φ = 60%. It is seen that the summation of exergy destruction in the pumps in Case 6 and Case 7 is less than that in the pump in Case 5. Figure 36 shows the variation of exergy destruction in the exhaust gas for all cases with respect to the change in pressure ratio at a fixed value of TIT = 1400 K and φ = 60%. As the pressure ratio increases, the turbine exhaust

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Fig. 32

Effect of variation of pressure ratio on exergy destruction in feedwater heater

Fig. 33

Exergy destruction in feedwater heater at TIT = 1400 K, rp = 14 and φ = 60%

Fig. 34

Effect of variation of pressure ratio on exergy destruction in pump

temperature decreases, so does the exergy destruction in exhaust gas for Cases 1 to 4. The exergy destruction in exhaust gas decreases for Cases 5 to 7 because the mass

flow rate of the gas decreases as the pressure ratio increases. Figure 37 shows the variation of exergy destruction in

Nikhil DEV et al. Performance analysis of combined cycle power plant

Fig. 35

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Exergy destruction in pump at TIT = 1400 K, rp = 14 and φ = 60%

is greater than that for Case 3 because the PT exhaust temperature for Case 4 is greater than the turbine exhaust temperature for Case 3. The exergy destruction in the exhaust gas for Cases 5 to 7 is the same. It is clearly observed that the premeditated installation of more components in the plant may decrease exergy destruction.

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Fig. 36 Effect of variation of pressure ratio on exergy destruction in exhaust gas

Conclusions

With the increase in pressure ratio, it is observed that the exergy destruction in the CC, HE, air filter, WHRB, ST, condenser, feedwater heater, pump and exhaust gas is decreased. While in the air compressor, GT, intercooler and REH, the exergy destruction is increased with the increase in pressure ratio. In this paper, seven cases are analyzed for the design parameters such as pressure ratio, etc. In the future, other operating parameters such as inlet air temperature may also be analyzed.

Notations $

Exergy rate/(kJ$s–1)

LHV

Lower heating value

R

Gas constant/(kJ$kg–1$K–1)

RH

Relative humidity

T

Absolute temperature/K

W

Work/(kJ$kg–1 (dry air))

cp

Specific heat at constant pressure/(kJ$kg–1$K–1)

cv

Specific heat at constant volume/(kJ$kg–1$K–1)

e

Specific exergy/(kJ$kg–1 (dry air))

h

Enthalpy/(kJ$kg–1 (dry air))

hf

Enthalpy of saturated water at process steam pressure

hg

Enthalpy of saturated vapor at process steam pressure

m

Mass/kg

E

Fig. 37 Exergy destruction in exhaust gas at TIT = 1400 K, rp = 14 and φ = 60%

the exhaust gas for all cases at a fixed value of TIT = 1400 K, rp = 14 and φ = 60%. The exergy destruction in the exhaust gas for Case 2 is greater than that for Case 1 because the mass flow rate of the gas for Case 2 is more than that for Case 1. The exergy destruction in the exhaust gas for Case 3 is less than that for Cases 1 and 2 because the mass flow rate of the gas is less than that for Cases 1 and 2. The exergy destruction in the exhaust gas for Case 4

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n

Number of moles

p

Pressure/bar

Qp

Process heat/(kJ$kg–1 (dry air))

rp

Compression ratio

s

Entropy/(kJ$kg–1$K–1)

t

Temperature/K

V

Specific volume/(m3$kg–1)

Greek symbols ω

Humidity ratio (kg of water vapor per kg of dry air)

φ

Relative humidity/%

ε

Effectiveness/%

η

Efficiency/%

g

Specific heat ratio

Subscripts AC

Air compressor

CC

Combustion chamber

D

Destruction

GT

Gas turbine

P

Product

Q

Heat

R

Regenerator

SG

Steam generator

W

Work

a

Ambient air

av

Average

f

Fuel

g

Gas

i

Inlet

l

Liquid

o

Outlet

sat

Saturated

v

Water vapor

w

Water

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