Exergetic evaluation of gas-turbine based combined cycle system with vapor absorption inlet cooling

Accepted Manuscript Research Paper Exergetic Evaluation of Gas-Turbine Based Combined Cycle System with Vapor Absorption Inlet Cooling Alok K. Mohapatra, Sanjay PII: DOI: Reference:

S1359-4311(18)31430-3 https://doi.org/10.1016/j.applthermaleng.2018.03.023 ATE 11911

To appear in:

Applied Thermal Engineering

Received Date: Revised Date: Accepted Date:

22 February 2016 8 December 2017 6 March 2018

Please cite this article as: A.K. Mohapatra, Sanjay, Exergetic Evaluation of Gas-Turbine Based Combined Cycle System with Vapor Absorption Inlet Cooling, Applied Thermal Engineering (2018), doi: https://doi.org/10.1016/ j.applthermaleng.2018.03.023

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Exergetic Evaluation of Gas-Turbine Based Combined Cycle System with Vapor Absorption Inlet Cooling

Alok K. Mohapatraa, Sanjay b* a

Department of Mechanical Engineering, Nalanda Institute of Technology, Bhubaneswar, India

b

Department of Mechanical Engineering, National Institute of Technology, Jamshedpur, India

* Corresponding author. Tel.: +916572373813; fax: +91 657-2373246.

1

E-mail address: [email protected].

Exergetic Evaluation of Gas-Turbine Based Combined Cycle System with Vapor Absorption Inlet Cooling

Abstract The aim of this study is to perform an exergetic evaluation of a gas-steam combined cycle power utility that is

integrated with vapor absorption inlet air cooling system. Effect of several thermodynamic

parameters on component-level and total exergy destruction are discussed. A detailed breakdown of exergy destruction in the cycle has been presented along with useful recommendations to improve exergy based cycle performance. The integration of inlet air cooling has been observed to increase exergy efficiency of combined cycle beyond a turbine inlet temperature of 1650 K. The combustor has been identified as the greatest source of irreversibility with highest exergetic improvement potential. The total exergy destruction for the cycle has been observed to reduce with increase in turbine inlet temperature (TIT) and decrease in compressor inlet temperature (TIC). At higher compressor pressure ratio (βcomp), except combustor all other components of topping cycle has been observed to suffer an increase in exergy destruction where as exergy destruction has been observed to reduce for all bottoming cycle components of bottoming cycle except the stack. The summation total of exergy destruction of all cycle components of the combined cycle plant is observed to be reduced at lower ambient relative humidity and lower ambient temperature. 2

Keywords: Combined cycle; Compressor inlet cooling; Exergy; Exergy Destruction Vapor Absorption;

NOMENCLATURE c

blade chord [m]

cp

constant pressure specific heat [kJkg -1 K -1 ]

f

factor [-]

F sa

correction factor to account for actual blade surface area [ -]

h

Specific Enthalpy [kJkg -1 ]

H r

lower heat value, heat of react ion / combustion [kJkg -1 ]

LHV Lower heat ing value [kJkg -1 ]  m

Mass flow rate [kgs -1 ]

p

Pressure [Pa]

Q

Heat [kJ]

R

Specific gas constant [kJkg -1 K -1 ]

s

Specific entropy [kJkg -1 K -1 ]

Sg

Blade perimeter at pitch line distance [m]

St in

Stanton number at stage inlet condit ion [ -]

t

Blade pitch [m]

T

Temperature [K]

T IC

Compressor inlet temperature [K]

T IT

Turbine inlet temperature [K]

v

Specific volume [m 3 /kg]

W

Power [kJ/s]

X

Exergy Flow [kJs -1 ] 3

X dest

Exergy destruction Rate [kJs -1 ]

SUBSCRIPTS a

Air

a,b,c,..n

Properties of various components of gas/ air mixture

b

Blade

c

compressor

ch

chemical

comb

combustor

cond

condenser

CV

Control volume

d/a

Deaerator

dest

destroyed

e

Exit

ex

exergy

f

Fuel

g

Gas

gt

Gas turbine

hrsg

Heat recovery steam generator

i

Inlet

j

Bleed point

mix

mixture

p

pump

ph

physical

s

Steam 4

st

Steam turbine

w

Water

GREEK SYMBOLS



Gas flow discharge angle [Deg]

β c o mp

Compressor pressure rat io

ε

Effect iveness [%]

E

Relat ive Exergy Consumpt ion ratio (%)

E

Fuel Exergy Deplet ion Rat io (%)

E

Productivit y Lack ratio (%)

ψ

Specific exergy [kJkg -1 ]



Specific humidit y



Relat ive humidit y [%]



Delta or difference

ACRONYMS B3PR

Basic 3 Pressure [for triple pressure with reheat combined cycle system without vapor absorption compressor inlet cooling]

BFP

Boiler Feed Pump

C

Compressor

CC

Combustion Chamber

CIC

Compressor Inlet Cooling

DA

Deaerator

Da

Dry Air

GT

Gas Turbine

IAC

Inlet Air Cooled/Cooling 5

IP

Improvement Potential [for exergy]

PLR

Productivity Lack Ratio

RECR

Relative Exergy Consumption Ratio

RH

Relative Humidity

VA3PR Vapor Absorption 3 Pressure [for triple pressure with reheat combined cycle system using vapor absorption compressor inlet cooling] VAIAC Vapor absorption inlet air cooling

Introduction Energy analysis is considered as a tool for evaluating sustainable performance of an thermal power systems. Improvement in energy efficiency can be said to reduce environmental of that power utility, impact owing to lower consumption of fossil fuel to generate equivalent amount of power leading to a greener environment. Hence, improvement in energy efficiency of power utility would ultimately lead to lowering impact of global warming, ozone depletion and finally environmental degradation. So, one of the major challenges to sustainable development lies in mitigating the environmental impact of energy use through increased resource utilization efficiency. Efforts to increase efficiency of power utilities could limit the environmental degradation significantly. Several methods have been reported to achieve, increase in efficiency of gas turbines and combined cycles [2-3]. Out of all these, increasing turbine inlet temperature is a well proven and promising technology [4-7]. Increasing turbine inlet temperature increases the combined cycle efficiency by raising the mean temperature of heat addition. However, with increase in turbine inlet temperature (TIT), the hot gas path blade coolant flow requirement increases due to larger heat flow from main working fluid to coolant stream. So increasing TIT has the adverse effect of increased irreversibility due to enhanced coolant flow rate and thus decreases available work output. Consequently, coupled with the need to improve efficiency through adoption of higher TIT , is the need to reduce increased irreversibility as a result of higher coolant flow. The irreversibility associated with the 6

higher coolant flow can be reduced drastically by reducing coolant temperature. Compressor inlet cooling fits this requirement. Implementation of compressor inlet cooling could significantly abate the negative impact of higher coolant requirement by increasing cooling effectiveness and thus reducing lost availability to produce work. There is insignificant published literature on the energy analysis of gas turbines and combined cycle plants subjected to both inlet air cooling and gas turbine blade cooling. Mohapatra et al. [8] have reported the first law based thermodynamic analysis of gas turbine and combined cycle plants with compressor inlet cooling and gas turbine blade cooling and concluded that evaporative compressor inlet cooling integrated to an internal convection cooled gas turbine cycle can improve the plant efficiency by 11.3% and plant specific work by upto 19.7%. The combined effect of air film blade cooling and vapor absorption compressor inlet cooling has been reported to increase the plant output and plant efficiency of gas turbine cycle by 18.3% and 7.48% respectively by Mohapatra and Sanjay [9]. For combined cycle, an increase of more than 7% in terms of plant output is reported. A comparative analysis of effect of compression and absorption inlet cooling on gas steam combined cycle is presented by Mohapatra and Sanjay [10]. However Though energy analysis is essential the benefits of energy systems evaluated only through energy analysis are not sufficient and therefore, it needs to be complemented by the exergy analysis. Exergy is defined as the maximum work obtainable by bringing a system into equilibrium with its environment. Every system not in equilibrium with its environment has certain quantity of exergy, while a system that is in equilibrium with its environment has, by definition, zero exergy since it has no ability to do work with respect to its environment. [11]

7

The unique feature of this article is the exergy analysis of a blade cooled gas turbine based combined cycle integrated with inlet air cooling system that has a significant scientific merit as cooled gas turbines are closer to real engines used by power utilities.

There is available literature on exergy based thermodynamic analysis of thermal energy systems. Bilgen [12] presented a second law based thermodynamic analysis of a gas turbine based cogeneration system and developed an algorithm for evaluating the thermodynamic performance and validated the simulated results with the reported data. Several alternatives to enhance the energy and exergy performance of combined cycle, including using dual-pressure heat recovery steam generator (HRSG), compressor inlet cooling and fuel gas preheating have been investigated by Sue and Chuang [13]. It is reported that a more precise prediction of plant efficiency can be made with the help of exergy analysis of the steam cycle system. The results obtained from the study have been used for engineering design and selection of components. The effects of steam injection ratios and feedstock mass flow rates on gas turbine performance integrated with steam injection and multiple effect evaporation has been discussed by Huang et al. [14]. The combustor and stack are reported to be the major sources of exergy destruction and exergy loss, respectively. Mahmoudi et al. [15] performed the energy and exergy analysis of simple and regenerative gas turbines combined with compressor inlet cooling using absorption refrigeration. They have reported that the first and second law efficiency values increase by around 2-7% for each 10 oC decrease in inlet air temperature. Khaliq and Dincer [16] have discussed the performance analysis of a combined heat and power system with inlet air cooling and evaporative aftercooling based on energy and exergy and reported combustor being the major source of irreversibility. The energy and exergy analyses of GT based combined cycle system with different turbine blade cooling technologies have been investigated by Sanjay et al. [17]. The component-wise exergy destruction 8

values were obtained as a function of GT cycle pressure ratio and turbine inlet temperature. The closed loop steam cooling is found superior than air film cooling and plant efficiency as high as 61.75% can be achieved using this technique. The summation of exergy loss of all components is also found to be lower for closed loop steam cooling with combustor being the largest source of exergy destruction. The second law based thermodynamic analysis of a cooled gas turbine has been performed by Khodak and Ramakhova [18]. Open-air cooling has been considered in their analysis. Exergy flow diagrams, detailing component-wise breakdown of exergy destruction, showed that about 5.3% of fuel exergy was destroyed due to adoption of turbine blade cooling related issues. There is available literature on gas turbines and combined cycle plants with a triple pressure level HRSG steam cycle. The exergy analysis of triple-pressure level steam cycle has been analyzed previously by Woudstra et al [19]. The value diagrams have been plotted to discuss about the effect of pressure levels on exergy loss. Combustion process is identified as the highest contributor to exergy destruction and the authors have suggested the integration of high temperature fuel cells like solid oxide fuel cells for further improvement in plant efficiency. The reduction in exergy loss due to heat transfer in HRSG and exergy of stack gas with increased pressure level has been identified as the prime cause of efficiency gain. 'An extensive review of energy and exergy analysis of thermal power plants has been performed by Kaushik et al [20]. Exergy analysis is used mainly to identify components with higher exergy destruction so that the design of such components can be improved to obtain improvement in plant performance. Once again combustion chamber (in case of gas fired thermal power plant) and boiler (in case of coal based thermal power plant) are identified as the primary source of exergy losses.

The available literature suggests that a more comprehensive exergy based thermodynamic analysis of gas turbines and combined cycle plants subjected to simultaneous integration of compressor inlet cooling and gas turbine blade cooling will be helpful in understanding areas of improvements to enhance their overall 9

performance. In particular, it will be of interest to evaluate absorption chiller technology in a combined cycle system as it allows effective utilization of energy resources by using waste low temperature energy. The integration of inlet air cooling enables a higher temperature to be adopted as compared to non-inlet cooled cycle and thereby ensures higher thermal efficiency [10]. Despite numerous researches on inlet air cooling the literature review reveals the following. 

All the studies made so far represent the first law thermodynamic analysis of combined cycle subjected to compressor inlet cooling.



Exergetic evaluation of inlet air cooled combined cycle has not been reported in literature.



The study of the effect of ambient temperature, ambient relative humidity, compressor pressure ratio, compressor inlet temperature and turbine inlet temperature on the exergetic performance evaluation of inlet cooling systems integrated to cooled gas turbine based combined cycle plant has not been investigated in literature and is one of the originalities of this work.

In addition to the above mentioned points, the available literature shows that effects of power augmentation approaches from exergy analysis perspective have not been fully examined to identify specific components of the combined cycle system contributing higher level of component level exergy destruction. Such an analysis will be helpful in identifying components with potential of improvement to enhance overall performance of the combined cycle system not achievable with commonly used energy balance approach. a systematic and comprehensive parametric exergy analysis of a combined cycle system, using f-class and advanced technology gas turbine based combined cycle, consisting of a threepressure HRSG with reheat and integrated with vapor absorption inlet air cooling has been reported in this research work

10

2.0

System Description

A GT based combined cycle system consisting of triple pressure level HRSG including reheating and integrated with vapor absorption inlet cooler has been analyzed in this study, a schematic of which is shown in Fig. 1. The main components of the combined cycle system considered in this study include a gas turbine consisting of an axial flow compressor, a combustor, and air filmed cooled turbine stages (as necessary), heat-recovery steam generator with reheat, steam turbine, condenser, deaerator, feed water pump, and the vapor absorption chiller unit. The enthalpy, exergy has been assumed to be zero at ambient conditions which is the reference (sink) temperature for the cycle. Table 1. lists the properties at different points in the combined cycle. Specific heat of all constituents individually is taken from the Y.S Touloukian and Makita Tadash [21] for the different temperature ranges and from that specific heat of combustion products are obtained from the standard theory of mixture as;



c p  f h * ya  c p ,a  yb  c p ,b  yc  c p ,c .......................  yn  c p ,n Where i = a, b, c …represents the constituents and



yi  mi / m (1a)

f h is the humidity correction factor and is represented as [10] f h  1  0.05h,e

(1b)

and  h,e is the relative humidity at the outlet of the vapor absorption cooler.

11

Table 1: VA3PR Combined Cycle thermodynamic Data

State

Pressure

Point

Location

Mass flow (kg/s)

Temperature (K)

(bar)

0

Ambient

1 (Air)

288

1.013

1

Compressor Inlet

1 (Air)

278

1.013

2

CC inlet

0.84 (Air)

716

23.31

3

GT Inlet

0.867 (Gas)

1750

22.84

4

GT Outlet

1.027 (Gas)

939

1.08

5

HPST inlet

0.11536 (Steam)

843

160

6

HPST outlet

0.11536 (Steam)

578

35

7

IPST inlet

0.13136 (Steam)

603

33.95

8

IPST outlet

0.13136 (Steam)

473

6

9

LPST inlet

0.16 (Steam)

473

6

10

LPST oultlet, X=0.88

0.16 (Steam)

305

0.05

11

Condenser inlet

0.16 (Steam)

305

0.05

12

Condenser outlet

0.16 (Water)

301

0.05

13

Deaerator inlet

0.16 (Water)

301

2

14

LPD inlet

0.02864 (Water)

369

6.6

15

IPD inlet

0.016 (Water)

369

37.35

16

HPD inlet

0.11536 (Water)

369

176

12

Thus, the enthalpy and entropy of gas is expressed as

h   c pg T dT

(2)

dT  R ln  p pa  T

(3)

T

Ta

s   c pg T   T

Ta

Equations 1-3 are applicable during calculations of stream properties throughout the cycle

13

VA3PR configuration boundary Vapor Absorption Compressor inlet cooling (VACIC) System

B3PR configuration boundary Heat to VA generator

Heat from HRSG

generator

Heat from HRSG to generator to Generator

20 Heat Exchanger

condenser

Heating Coil

19

Condensing Water Coil

Stack

Expansion Valve

17

18

HRSG

21 16

IPD

HPD

LPD

0 evaporator

evaporator

absorber

Economiser

Ambient air

Refrigerant (vapor)

GT Exhaus t

1

Cooled air

Cooling Coil m f

4

CC

2

C

3

7’’ GT Exhaust to HRSG

8

14

7’ 6

5

GT

8’

HPST

7

10

9

IPST

15

LPST LP, BF Pump

Alternator

IP,BF Pump

11 Condenser Cooling water

‘HP’,BF Pump

Coolant A ir (Film Cooling)

Deaerator

Conde nser

12

Make-up water CEP

13

14 Figure 1. Schematic of 3PR combined-cycle configuration with VA inlet air cooling (VA3PR) and without inlet air cooling (B3PR)

3.0

PERFORMANCE PARAMETERS

The various performance parameters considered during exergy analysis of the considered combined cycle system include the following: 1. Exergy Efficiency: It is defined as the ratio of the exergy of products (exergy output) to the exergy of supply (exergy input).

 ex 

2.

Xe Xi

(4)

Fuel Exergy Depletion Ratio: It is defined as the ratio of the component exergy destruction to

E 

the fuel exergy input [22]

X dest X fuel

(5)

3. Relative Exergy Consumption Ratio: The Relative Exergy Consumption Ratio (RECR) of a combined cycle system’s component is calculated as the ratio of the component’s exergy destruction to the total exergy destruction of the system [23].

E 

X dest,component

(6)

X dest,total

4. Exergetic Improvement potential: This is an exergy based performance parameter calculated by the following equation [24,25].

IPex  1  ex    X i  X e  5. Productivity Lack ratio (  E ): Productivity Lack Ratio (PLR) is an exergy based performance parameter and is calculated by the ratio of the component’s exergy destruction to the product exergy [23].

E 

X dest, component X product

15

(7)

4.0

RESULTS AND DISCUSSION

A computer program “X-Comb”, developed in-house using C++ which has been previouslytested and authenticated by the authors [9, 26] while analyzing other power plant systems in previous studies, and has been used for evaluating exergertic performance parameters. A detailed parametric simulation has been performed for two values of TIT (1555 K and 1750 K) and two values of comp (16 and 23) to identify the impact of using vapor absorption inlet cooling system on exergy related performance parameters. The results of parametric analysis are presented and discussed in details in various subsections below. It must be mentioned that TIT values are selected to represent GT technology of the 1990’s (1300 0C or less, typical of F-class machines) and more advanced technology machines (1500 0

C G class machines) which became operational in early 2000s.

The gas-turbines are designed to operate at ISO condition (T a = 15o C and RHa = 0.6). But in the present article for analysis, the compressor inlet temperature and ambient relative humidity have been varied to simulate seasonal changes in ambient conditions and hence ISO condition is not achieved at all times. However, the results obtained in this article have been validated with characteristics of MS 7001 gas-turbine from General Electric [27,28] under off design conditions and with compressor inlet cooling and they have been found to compare well. The results obtained for combined cycle compares well with the results presented by Dechamps [4], Bolland [5] and GE S207FB combined cycle plant [27]. The results obtained for a cooled gas turbine by means of air film cooling have been validated against the results of El-Masri [29] and shows a good comparison with an acceptable variation of 2.5%–3.5% (for coolant mass flow rate). The results of inlet air-cooling model are validated by comparing the results with those presented in Boonassa et al. [30], Wang and Chiou [31], and PG 7241 FA model of inlet air-cooled CC plant by GE power [28]. 16

The exergy analysis results have been validated against the results presented in [26,32,33] and is found to compare well. To provide a performance comparison, the simulation is performed utilizing the same design parameters selected in the published results and existing CC plants. The results show good comparison with the results of published papers and existing CC plants.

4.1

Effect

of

ambient

relative

humidity

and

ambient

temperature

on

plant

specific work and plant efficiency of 3PR configuration with and without VAIAC. Fig. 2(a) shows the effect of ambient conditions on the plant specific work for 2PR configuration with and without VA inlet cooling. It can be observed that the plant specific work of VA3PR cycle is higher than the B3PR cycle at ambient relative humidity of 0.2 and ambient temperatures of 293 K and 313 K respectively. B 3P R C onfiguration V A 3P R C onfiguration

r p ,c = 23,

T IT = 1750 K ,

C IT = 278K

P la n t sp e cific w o rk (kJ/kg )

640

620

600

580

560

540 R H a = 0 .2 ,T a = 2 9 3 K

R H a = 0 .2 ,T a = 3 1 3 K

R H a = 0 .6 ,T a = 3 1 3 K

A m b ie n t co n d itio n s F ig . 2 (a ): E ffe ct o f a m b ie n t re la tiv e h u m id ity a n d a m b ie n t te m p e ra tu re o n p la n t sp e cific w o rk fo r 3 P R co n fig u ra tio n w ith a n d w ith o u t V A IA C

17

This is because, at these values of ambient temperature and ambient RH, the performance gain due to increase in gas cycle work (due to reduced work of compression and lesser blade coolant requirement) and decrease in steam cycle work (due to heat energy extracted from HRSG to run the refrigeration system) is always positive. This net gain is higher at higher ambient temperature because of greater difference between dry bulb temperature (DBT) and wet bulb temperature (WBT) resulting in more effective inlet air-cooling. This gain, however turns negative when the ambient RH increases to 0.6 and so the plant specific work of VA3PR configuration is lower than that of B2PR configuration.

B 3P R C onfiguration V A 3P R C onfiguration

r p ,c = 23,

T IT = 1750 K ,

C IT = 278K

58

P la n t e fficie n cy (% )

56

54

52

50

48 R H a = 0 .2 ,T a = 2 9 3 K

R H a = 0 .2 ,T a = 3 1 3 K

R H a = 0 .6 ,T a = 3 1 3 K

A m b ie n t co n d itio n s F ig . 2 (b ): E ffe ct o f a m b ie n t re la tiv e h u m id ity a n d a m b ie n t te m p e ra tu re o n p la n t e fficie n c y fo r 3 P R c o n fig u ra tio n w ith a n d w ith o u t V A IA C

Fig. 2(b) shows the effect of ambient conditions on the plant efficiency for 3PR configuration with and without VA inlet cooling. It can be observed that the addition of VA inlet air cooling improves the

18

efficiency of 3PR cycle for entire range of ambient condition. The increase in efficiency is observed to be more pronounced at lower ambient temperature. This can be explained as follows: At lower ambient temperature the relative gain in specific work due to addition of inlet air cooling is reduced as explained in Fig 1(a). However at lower ambient temperature, the fuel energy input for B3PR configuration increases due to higher difference between TIT and TIC (TIC and Tamb are same for non inlet cooled configuration) where as that for VA3PR configuration remains unchanged (TIC is maintained constant at 278 K).

The integration of inlet air cooling is observed to improve plant

efficiency of 3PR combined cycle by 4.5 % at an ambient temperature of 293 K against 1.2 % at Tamb= 313 K.

4.2

Effects of TIT on the overall exergy efficiency of the combined cycle system

To start with, we first present the analysis (Fig. 2) of effect of changing TIT for a fixed value of compressor pressure ratio (comp = 23), keeping compressor inlet temperature (TIC = 293 K) and ambient condition (Tamb = 313 K, RHamb = 20%) constant for cycle with and without vapor absorption cooling method.

19

5 5 .5 5 5 .0

V A 3P R B 3P R

O v e ra ll E x e rg y E ffic ie n c y (% )

5 4 .5 5 4 .0 5 3 .5 5 3 .0 5 2 .5 5 2 .0 5 1 .5 5 1 .0

 co m p = 23, T a m b = 313 K ,

5 0 .5

R H a m b = 0.2, T IC = 293 K for V A 3P R cycle

5 0 .0 1400 1450 1500 1550 1600 1650 1700 1750 1800 1850 1900 1950 2000

T urbine inlet tem perature (K ) F ig.3 . V ariation of E xergy efficiency w ith T i,T for 3P R cycle w ith and w ithout inlet air cooling

The optimum plant efficiency in case of B3PR cycle is observed to be at a Ti,T = 1675 K. This is because beyond a Ti,T of 1675 K, the efficiency benefits of higher temperatures may be more than offset by the increased losses associated with the enhanced blade coolant flow rate. This optimum temperature corresponding to maximum efficiency is increased to 1750 K for IAC3PR configurations. This is because for inlet air cooled configurations the Ti,C reduces, resulting in a subsequent reduction in temperature of coolant bled from the compressor. This ensures a higher Ti,T to be adopted for inlet air cooled cycles compared to non-inlet cooled cycle and therefore higher exergy efficiency is achieved. At higher temperature the heat to work conversion is higher resulting in higher exergy efficiency of inlet air cooled configuration. The histogram in Fig. 3 shows the rational efficiency of gas turbine and steam turbine along with summation of exergy destruction in all components for 3PR combined cycle with and without inlet air cooling. It is observed that the rational efficiency of GT in the inlet air cooled configuration is higher. This is because as the inlet air is cooled, both the gas turbine net-work(due to lower compression 20

work) and fuel energy input increases. The combined effect of these two factors is such that the rational efficiency of inlet air cooled configuration is higher than that of B3PR configuration. The rational efficiency of steam turbine however decreases for VA3PR configuration due to the combined effect of reduction in exergy of product (steam turbine work) and increase in exergy of

E xe rg y E fficie cy (% ) & F u e l E xe rg y D e p le tio n R a tio (% )

supply (fuel exergy input).

com p = 2 3 , T am b = 3 1 3 K , T IT = 1 7 5 0 K R H am b = 0 .2 , T IC = 2 9 3 K fo r V A 3 P R c yc le

55

B 3 P R C ycle V A 3 P R C ycle

P la n t E xe rg y E ffic ie n c y (% )

F u e l E xe rg y d e p le tio n R a tio (% )

50 45 40

C o m p o n e n t E xe rg y e ffic ie n c y (% )

35 30 25 20 15 GT

ST

C yc le

C yc le

F ig 4 . E x e rg y e fficie n cy a n d F u e l E x e rg y D e p le tio n R a tio o f B 3 P R a n d V A 3 P R cycle .

4.3

Effects of vapor absorption inlet-air-cooling on exergy destruction of combined cycle system’s

main components Figures 5(a) and 5(b) will help in understanding reasons to a certain extent for the observed reduced overall exergy efficiency at TIT < 1675 K with implementation of the vapor absorption inlet air cooling as seen in Fig. 2. The exergy destruction values at T IT of 1550 K, expressed as a percent of fuel exergy input, are significant only for topping cycle components (compressor, combustion chamber and turbine) and are very small (< 1 %) for the bottoming cycle components (see Fig. 4a). 21

V A 3P R cycle, T a m b = 313K ,R H a m b = 0.2,

50

T = 293 K for V A 3P R cycle, IC

co m p = 23

V A 3P R cycle B 3P R cycle

40 35 30 25

T

IT

=1550 K

8 6 4

T o ta l

DA

BFP

C ond.

ST

S ta c k

HRSG

CC

C IC

0

GT

2

C

F u e l E x e rg y D e p le tio n R a tio (% )

45

C o m b in e d cycle co m p o n e n ts F ig 5 (a ). C o m p a riso n o f F u e l E x e rg y D e p le tio n R a tio fo r 3 P R cycle w ith a n d w ith o u t v a p o r a b so rp tio n in le t co o lin g a t T IT = 1 5 5 0 K

50

V A 3P R cycle, T a m b = 313K ,R H a m b = 0.2, T = 293 K for V A 3P R cycle,

co m p = 23

45 40

V A 3P R cycle B 3P R cycl

35 30 25

T

IT

=1750 K

8 6 4

T o ta l

DA

BFP

C ond.

ST

S ta c k

HRSG

GT

CC

0

C

2 C IC

F u e l E x e rg y D e p le tio n R a tio (% )

IC

C o m b in e d cycle co m p o n e n ts F ig 5 (b ). C o m p a riso n o f F u e l E x e rg y D e p le tio n R a tio fo r 3 P R cycle w ith a n d w ith o u t v a p o r a b so rp tio n in le t co o lin g a t T IT = 1 7 5 0 K

Also, among the identified topping cycle components, combustor has a major share (~ 30% or more) of the exergy destruction and, thus is the biggest contributor to the overall irreversibility of the combined cycle system. Furthermore, the observed increase (even though very small) in the amount of exergy destruction in presence of the vapor absorption inlet air cooling for the topping cycle components can be attributed to larger internal losses in these components due to higher than optimum 22

pressure ratio corresponding to TIT of 1550 K. It is well known, as discussed by Bhargava et al [34], from the gas turbine cycle performance characteristics that an optimum pressure ratio to achieve maximum specific work reduces as the TIT value decreases. Also, as discussed by Horlock [35], combined cycle efficiency is influenced by the optimum overall cycle pressure ratio which decreases with reduced turbine inlet temperature. 4.4

Effects of TIT and βcomp on exergy destruction of combined cycle system’s main components

An examination of exergy destruction of key components of the combined cycle reveals some interesting results which were not clear while considering the overall exergy levels as discussed earlier in Fig. 4. Fig 6 (a), exhibiting the effect of TIT on exergy destruction of components combined with the selected inlet air cooling method shows that the highest exergy destruction ratio (more than 30%) has been observed in the combustor, and this value is lower at higher TIT . This relates to the discussion by Leites, Sama and Lior [36] in their work in section 4.2, wherein they have suggested methods to reduce exergy destruction: “heat of reaction to have a positive thermal exergy, the reaction temperature, T, must be greater than To.... Thus, there must be a reaction temperature, greater than To, but less than Teq (reaction equilibrium temperature), where the work effect is optimal and; ” therefore, where the exergy destruction reduces and in conclusion recommend that the increase in the value

of

TIT

should

be

such

that

exergy

23

destruction

in

combustor

reduces.

45

V A 3P R cycle, T a m b = 313K ,R H a m b = 0.2, T = 293 K , IC

co m p = 23

F u e l E xe rg y D e p le tio n R a tio (% )

40 T 35

T

IT IT

= 1550 K = 1750 K

30 25 8 6 4 2

T o ta l

DA

BFP

C ond.

ST

S ta c k

HRSG

GT

CC

C

C IC

0

C o m b in e d cycle co m p o n e n ts F ig 6 (a ). F u e l E x e rg y D e p le tio n R a tio o f V A 3 P R cycle fo r v a ria tio n in T IT

As TIT value increases, exergy destruction ratio in the gas turbine increases because of increased flow of coolant air needed for hot gas path components (combustor and turbine section blades) and associated losses and that of steam turbine increases due to increase in the amount of steam generated and entering the bottoming cycle. The exergy destruction ratio in the HRSG increases because of higher amount of steam handled and the increased temperature difference between the two heat exchanging fluids. The overall exergy destruction ratio has been observed to be lower at higher value of TIT.

24

45

V A 3 P R cycle , T a m b = 3 1 3 K ,R H a m b = 0 .2 , T

I,T

F u e l E xe rg y D e p le tio n R a tio (% )

40

= 1 7 5 0 K , T I,C = 2 9 3 K

 co m p = 1 6  co m p = 2 3

35 30 25 8 6 4 2

T o ta l

DA

BFP

C ond.

ST

S ta c k

HRSG

GT

CC

C

C IC

0

C o m b in e d c yc le c o m p o n e n ts F ig 6 (b ). F u e l E x e rg y D e p le tio n R a tio o f V A 3 P R c yc le fo r v a ria tio n in c om p

Fig. 6 (b) shows that the exergy destruction ratio for the compressor, at a fixed value of TIT (1750 K) increased with increase in compressor pressure ratio (βcomp) mainly because of increased compression work (unproductive use of available energy). With increase in βcomp, the exergy destruction ratio for the combustor decreases mainly due to the increase in the compressor discharge temperature resulting in increased air temperature entering the combustor and the corresponding reduction in specific fuel consumption. At higher value of βcomp, a higher level of exergy destruction ratio is seen with GT, primarily due to higher expansion ratio and also due to greater cooling air required, and mixing irreversibility resulting from higher coolant air requirement. It can also be noted that the exergy destruction ratio of the topping cycle components increases whereas, that of bottoming cycle components decreases with increase in the value of βcomp. This is mainly because at higher βcomp, except combustor all other components of topping cycle suffer an increase in exergy destruction ratio, whereas exergy destruction ratio reduces for all components of the bottoming cycle except the stack. The sum total of exergy destruction ratio of all components of the combined cycle plant is found lower 25

at a higher value of compressor pressure ratio, implying that vapor absorption inlet air cooling will be energy efficient for the selected conditions. 4.5

Effects of TIC on exergy destruction of combined cycle system’s components

As evident from Fig 7, the overall exergy destruction reduces with decrease in the value of T IC at a fixed value of the other design and operating conditions (TIT, comp, Tamb, and RHamb).

45

V A 3 P R c yc le ,  c o m p = 2 3 , R H a m b = 0 .2 , T

F u e l E xe rg y D e p le tio n R a tio (% )

40

IT

= 1 7 5 0 K ,T am b= 3 1 3 K T

35

T

i,c i,c

=303 K =293 K

30 25 8 6 4 2

T otal

DA

BFP

C ond.

ST

S tack

HRSG

GT

CC

C

C IC

0

C o m b in e d c yc le c o m p o n e n ts F ig 7 . F u e l E x e rg y D e p le tio n R a tio o f V A 3 P R c yc le fo r v a ria tio n in T IC

This is mainly due to an increased cooling load, as is evident from increased exergy destruction level of the vapor absorption cooling system (represented as CIC in Fig. 6), and effective utilization of waste exhaust energy from the gas turbine. Effective utilization of waste exhaust energy is also evident from the reduced exergy destruction for the HRSG unit (see Fig. 6). The effect of reducing TIC has a significant effect on reducing exergy destruction of compressor and gas turbine. This is because of increased compressor power and reduction in coolant bled from compressor and subsequently 26

supplied to the gas turbine. Another factor contributing to the overall reduced exergy destruction is an increased power output of the gas turbine (indicated by its reduced exergy destruction). The exergy destruction ratio in combustor slightly increases with decrease in TIC value due to higher fuel consumption at a fixed value of TIT. 4.6

Effect on component and overall exergy destruction ratio of inlet air cooled 3PR combined

cycles for variation in Tamb and RHa The major effect of change in ambient temperature, for a fixed value of T IC and other design and operating conditions, is expected on exergy destruction of vapor absorption system, HRSG and stack as can be seen in Fig. 8a.

45

V A 3 P R c yc le ,  co m p = 2 3 ,R H a m b = 0 .2 , T =1750 K , T

F u e l E xe rg y D e p le tio n R a tio (% )

40

IT

IC

=293 K

T am b = 3 2 3 K T am b = 3 0 3 K

35 30 25 8 6 4 2 0 -2

T otal

DA

BFP

C ond.

ST

S tac k

HRSG

GT

CC

C

C IC

-4

C o m b in e d c yc le c o m p o n e n ts F ig 8 (a ). F u e l E x e rg y D e p le tio n R a tio o f V A 3 P R c yc le fo r v a ria tio n in T am b

As a result, the overall exergy destruction is noted to increase at high ambient temperature. It is observed that the exergy destruction in the combustor is highest and it exhibits a lower value at lower 27

ambient temperature. This relates to the same discussion in Leites, Sama and Lior [36] wherein they have pointed out that the maximum work that could be generated from an irreversible oxidation process such as combustion is the work potential of the heat at the reaction temperature relative to the ambient temperature. So at a lower ambient temperature, T a the work effect is increased and the exergy destruction is reduced. It must be noted that exergy destruction levels for the aforesaid three components are small but relative changes are large. Exergy destruction of HRSG and stack decreases with increase in ambient temperature because of effective waste energy utilization and less waste energy in HRSG and stack, respectively. Whereas, exergy destruction of vapor absorption system increases because of the increased cooling load as represented by increased exergy destruction level of CIC (see Fig. 8a) at high ambient temperature. The effects of ambient relative humidity (RHamb) on the combined cycle system’s components and the overall exergy destruction ratio for VA3PR cycle are shown in Fig 8 (b).

28

45

V A 3 P R cycle ,  co m p = 2 3 ,T a m b = 3 2 3 K , T =1750 K , T IT

F u e l E xe rg y D e p le tio n R a tio (% )

40

IC

=293 K

R H am b = 0 .6 R H am b = 0 .2

35 30 25 8 6 4 2

T o ta l

DA

BFP

C ond.

ST

S ta c k

HRSG

GT

CC

C

C IC

0

C o m b in e d c yc le c o m p o n e n ts F ig 8 (b ). F u e l E x e rg y D e p le tio n R a tio o f V A 3 P R c yc le fo r v a ria tio n in R H am b

As expected, main component impacted due to change in relative humidity is the vapor absorption system for which cooling load increases relatively significantly with the increase in ambient relative humidity level. This results in increased exergy destruction for this component as shown in Fig. 8 (b). With decrease in RH, though the steam turbine specific work increases, the exergy destruction is also increased due to higher mass of steam handled. The sum total of exergy destruction ratio of all components of the combined cycle plant has been observed higher at an increased value of the ambient relative humidity of 60% (RHamb = 0.6). 4.7

Exergetic Improvement potential of inlet-air-cooled 3PR combined cycle components

29

1 0 6 .1 7

E xe rg e tic Im p ro ve m e n t P o te n tia l (kW )

105

com p = 2 3 , T am b = 3 1 3 K , T IT = 1 7 5 0 K

90

R H am b = 0 .2 , T IC = 2 9 3 K fo r V A 3 P R c yc le

75

60

45

30

15

0

4 .0 7

1 .5 9 C IC

C

6 .9 8

4 .2 2

4 .1 6

CC

GT

1 .9 6 HRSG

ST

C ond.

C o m b in e d C ycle C o m p o n e n ts F ig . 9 . E x e rg e tic Im p ro v e m e n t P o te n tia l o f m a jo r V A 3 P R co m b in e d cycle co m p o n e n ts

An important exergy based parameter, exergetic improvement potential (IP ex) of analysed combined cycle components have been presented in Fig. 9. Van gool [24] has initially used the term “Exergetic Improvement Potential” when analyzing different processes or sectors of economy which later has been referred by Hammond and Stapleton [25] to describe the exergy analysis of energy systems.

In order to achieve improvement in exergy efficiency, the exergy destruction needs to be minimized. A component with higher exergy destruction will obviously have lower exergy efficiency (both of which tends to increase IP). So “Improvement Potential” is a measure of a component’s potential for improvement that will result in lower exergy destruction and consequently higher exergy efficiency. This is used to identify components so that the one with higher IP may preferably be chosen for design improvements in order to improve exergy performance of the system.

30

Combustor is identified as the component with highest exergetic improvement potential of 106.17 kW, which implies that combustion related improvements could significantly improve the exergetic performance of a combined cycle. This can be achieved by employing fuel cells or using metal oxides in combustion process. Hence the high degree of irreversibility associated with the combustion process can only be improved by replacing the combustor with alternative technique such as Chemical Looping Combustion (CLC) [36]. The exogenous exergy destruction in combustion chamber can be reduced through changes in the performance of the remaining plant components and in the structure of the overall system [37] .The inlet air cooler is observed to have the lowest IPex with a value of 1.59 kW followed by steam turbine with 1.96 kW. 4.8

Relative Exergy Consumption Ratio and Productivity Lack Ratio of major VA3PR combined

cycle Components. Fig. 10 represents the relative exergy consumption ratio and productivity lack ratio of major VA3PR cycle components at TIT of 1750 K and other fixed design and operating conditions. As noted earlier, the combustor section of the GT has been identified as the greatest source of relative exergy consumption followed by the turbine and compressor sections. This is mainly due to the fact that irreversibility associated with the combustion reaction and the heat transfer across a large difference in temperature between the primary zone combustion gases and mixing of the working fluids. Combustion processes exhibit very high thermodynamic inefficiencies caused by chemical reaction, heat transfer, friction and mixing [37] which are not identified through First law analysis of energy balance. The major contribution to the exergy destruction in typical gaseous hydrocarbon fuel consumption is due to the internal thermal energy storage between particles within the system. The relative exergy consumption ratio of all the other components lies in the range of 0.5 to 4 %.

31

R elative E xergy C onsum ption R atio (% ) and P roductivity Lack R atio (% )

70 R e la tiv e E x e rg y C o n s u m p tio n R a tio (% ) P ro d u c tiv ity L a c k R a tio (% )

65

c o m p = 2 3 , T a m b = 3 1 3 K , T IT = 1 7 5 0 K

60

R H a m b = 0 .2 , T IC = 2 9 3 K fo r V A 3 P R c yc le 55 50

15 10 5 0 C IC

C

CC GT HRSG M a jo r c o m b in e d c yc le c o m p o n e n ts

ST

C ond.

F ig 1 0 . R e la tiv e E xe rg y C o n s u m p tio n R a tio a n d P ro d u c tiv ity L a c k R a tio o f m a jo r V A 3 P R c o m b in e d c yc le c o m p o n e n ts

One more term “Productivity Lack Ratio” has been plotted along with relative exergy consumption ratio [23]. Productivity lack provides the product loss in the form of exergy destruction or quantifies the product exergy potential, lost due to exergy destructions. Identical to the trends observed for relative exergy consumption ratio the productivity lack ratio is highest for combustion chamber followed by gas turbine and compressor. For all other components its value lies between 0.4 to 3 %.

Conclusions

Based on the systematic and comprehensive parametric exergy analysis of a combined cycle system consisting of three-pressure HRSG with reheat integrated with vapor absorption cooling, the following concluding remarks can be made: 

The integration of vapor absorption inlet air cooling system has been observed to enhance energy and exergy efficiency of 3PR combined cycle .



Combustor has been identified as the greatest source of irreversibility (>60 %) followed by gas turbine and compressor. Therefore, combustor related improvements can be expected to significantly improve the exergetic performance of gas-steam combined cycle. 32



For a given design conditions (values of T IT, βcomp and TIC) and the operating conditions, the exergy destruction of the topping cycle components including inlet air cooling system increases with increase in the ambient temperature as a result of implementation of the vapor absorption inlet air cooling system mainly due to reduced inlet air cooling effects or increased cooling load.



Combustor and inlet air cooler have been reported as having highest and lowest exergetic improvement potential with 106.17 kW and 1.59 kW respectively.



The total exergy destruction has been observed to be lower at higher value of TIT largely due to the reduction in combustor irreversibility.



Combustor, Turbine and Compressor are among the prime contributors to exergy destruction and product exergy loss.



At higher βcomp, except combustor all other components of topping cycle has been observed to suffer an increase in exergy destruction where as exergy destruction has been observed to reduce for all components of bottoming cycle except the stack. The sum total of exergy destruction of all components of the combined cycle plant is lower at higher value of compressor pressure ratio.



The sum total of exergy destruction of all components of the combined cycle plant is higher at higher value of ambient temperature the major contribution being from bottoming cycle components.



The ambient relative humidity only marginally affects the exergy destruction of all components except absorption inlet cooler.

Acknowledgement The authors are highly obliged to Dr. Rakesh K. Bhargava, Founder & President of Innovative Turbomachinery Technologies Corp for his expert advice and guidance during the entire process of writing of this research publication His expertise in the field of turbo-machinery design and consultancy is reflected in the article.

33

Appendix A Exergy Based Modeling and Governing Equations In the absence of nuclear, magnetic, electrical, and surface tension effects, the total exergy of a system can be divided into four components: physical exergy, kinetic exergy, potential exergy, and chemical exergy [1]. In this work, kinetic exergy and potential exergy have been assumed to be negligible as the elevation of various components of the combined cycle system and flow velocities through them have negligible changes. Mass, energy, and exergy balances for any control volume at steady state of operation can be expressed, respectively, by

 m   m

(A1)

 e he   m  i hi Q  WCV   m

(A2)

i

e

 e e   m  i i  X dest X heat  X W   m

Where, X heat =



(A3)

To    Qi and X W = W are exergy associated with flow of heat and work i 

 1  T 

respectively,  is the specific exergy and X dest is the exergy destruction The total exergy is given as X  X ph  X ch

(A4)

Where, physical exergy is given by,   ph  m  h  ho  To s  so  X ph  m

(A5)

and chemical exergy of mixture is given by,

34

n  n  ch ch   mix   xi chi  RTo  xi ln xi  X mix m m i 1  i 1 

(A6)

An approximate formulation for chemical exergy of gaseous hydrocarbon fuel is given as [38]

 

 f  LHV f  1.033  0.0169

H 1  0.0698  C C

(A7)

In the present work, moist air which may be considered as a mixture of ideal gases, dry air, and water vapor is analyzed. The exergy of moist air per unit mass of dry air is represented by the following equation [39]:

T

T 

  (c p  c p )Ta   1  ln   1  1.608 RaTa ln  T T p a  a  a a

p

v

  1.608  1  1.608 a  RaTa 1  1.608  ln    1.608 ln   1  1.608   1.608 a 

   (A8)

Expressions of Exergy destruction rate for each component of system in terms of state points is presented in Table 3.

Table 3. Expressions of Exergy destruction rate for each component of system Component

Exergy Destruction Rate Expression

Inlet Air Cooler

X 19  X 1  X 20  X dest

Compressor

X 1  Wc  X 2  X dest  X coolant

35

Combustor

X 2  X f  X 3  X dest

Cooled Gas Turbine

X 3  X coolant  X 4  X dest  Wgt

HRSG

X 4  ( X16  X17  X18 )  X19  X 21  ( X 5  X 7''  X 8' )  X dest

Steam Turbine

X 5  X 7  X 9  Wst  ( X 6  X 8  X10  X11)  X dest

Condenser

X11  X12  X dest  X Q

Deaerator

X 10  X 14  X 15  X dest

Pump

X 12  Wpump  X 13  X dest

The detailed mass, energy and exergy balance of various combined cycle components can be expressed with help of the following equation: (i) Vapor Absorption Inlet Air Cooler

In this method of cooling, inlet air to the compressor is cooled from ambient temperature to a lower temperature by means of an ‘ammonia-water’ vapor absorption refrigeration system. Of the following three options of thermal energy source to drive the system, option-3 has been found suitable, for the reason discussed as under: 1. Option-1: Exhaust gases at the exit of HRSG, has the limitation of its temperature close to the dew point temperature and further extraction of heat from it would lead to condensation of moisture in the flue gases having SOx/NOx. This results in weak acid formation on the heat exchanger tubes of absorption chiller and resulting its corrosion. 2. Option-2: Utilizing steam extracted from IP/LP steam turbine casing having energy just sufficient to meet the absorption chiller energy requirement. This is not appropriate as it results 36

in greater loss of steam cycle work due to bleeding of steam that would otherwise have been used to generate steam cycle work. 3. Option-3: Exhaust gas extracted at an appropriate point prior to the exit of HRSG having sufficient heat energy, has been found to be most suitable for utilization in absorption chiller. The heat from HRSG is extracted at an exhaust gas temperature of 120oC which is the ideal generator temperature in an aqua ammonia vapor absorption system [31]. To model the vapor absorption system the following assumptions have been made: 

The temperature of gas leaving the vapor absorption (VA) generator is restricted to a minimum level of 373K to safeguard against the possible moisture condensation present inside the gas.



The concept of heat transfer effectiveness of VA generator (εgen) has been introduced to account for the actual heat input in the VA generator.



The pump work of the absorption system is negligible. The amount of heat required to run the absorption refrigeration system for the selected COP has been determined by:

Qabs  QCL COPVA

(A9)

Hence, the heating load of the VA generator, taking into account its effectiveness, is determined by:

Qgen  Qabs  gen

(A10)

The mass and exergy balance of VA inlet air cooling can be expressed as  a ,i  m  a ,e m

and

X dest,VACIC  ma  a ,i  a ,e   ms  s ,i  s ,e  

 s ,i  m  s ,e m

(A11)



37

(A12)

(ii) Compressor  c ,i  m  c.e   m  coolant, j m

(A13)

 c,e hc,e   m  coolant, j hcoolant, j  m  c,i hc,i  Wc  m

(A14)

X dest,c  Wc  m c, i c, i   m coolant , j  coolant , j  m c, e c, e

(A15)

(iii) Combustor

m e comb  m i  m f comb  f  H r  comb  m  e  he  m  i  hi comb m

(A16) (A17)

The physical exergy of natural gas is calculated by Eq. ( A4) and the chemical exergy of natural gas is calculated as follows [38, 40]

 H 1   X ch  m f  f  m f   LHV f  1.033  0.0169  0.0698  f C C   

X CC  X fph  X ch f f



  X dest,comb  X CC f  mg ,i  g ,i  mg ,e . g ,e

(A18)

(A19)



(A20)

comb

(iv) Cooled Gas Turbine In this work, the gas turbine blades have been modeled to be cooled by air-film cooling (AFC) method. The cooling model used for cooled turbine is the refined version of that used by Louis et al. [26]. The mass flow rate of coolant required in a blade row is expressed as [26]:

a coolant 

m coolant  St in  c p , g   S g  Fsa   Tg ,i  Tb      m g    c p ,coolant   t  cos  mean  Tb  Tcoolant,i 

(A21)

Where, Sg  2c, Sg/tcos(mean) = 3.0, Fs,a = 1.05,  mean = 45 (for stator),  mean = 48 (for rotor), and Stin = 0.005. 38



  m

 g ,i  ( h g ,i  h g , e )  Wgt  m

coolant

 (hcoolant,i  hcoolant,e )



X dest, gt  m g ,i  g ,i   m coolant coolant,i  coolant,e   m g ,e  g ,e  Wgt

(A22) (A23)

(v) HRSG  g , i  g , i  m  g , e  g , e    m  s , i  s , i  m  s , e  s , e  X dest,hrsg  m

(A24)

(vi) Steam turbine

Wst 

 m  h

stages

s .i

s ,i

 hs , e 

hs ,i  hs ,e   st hs ,i  hs ,e isentropic

(A25)

(A26)

The exergy destroyed in HP, IP and LP stages have been obtained from the following equation

X dest, st 

stages

 m

s, i

 s , e  s , e  Wst  s , i  m

(A27)

Where, m s.i  amount of steam entering to the respective main turbine stages as per configuration. (vii)

Condenser

m w,i  m w, e

 s ,i  m  cond,e m

and

 s ,i .hs ,i  m  cond,e .hcond,e )  m  w.(hw,e  hw,i ) (m

(A28)

(A29)

 s ,i . s ,i  m  cond, e . cond,e )  (m  w,i w,i  m  w,e w,e ) X dest,cond  (m (viii)

De-aerator

 s ,d / a,i  m  cond  m  s ,d / a,i   m  s ,d / a,e m  s ,d / a,i .hs ,d / a,i  m  cond  m  s ,d / a,i .hs ,cond,e  m  s ,d / a,e .hs ,cond,e m 39

(A30)

(A31)

X dest, d / a  m s,d / a,i  s,d / a,,i  m cond  m s,d / a,i   s,cond,e  m s,d / a,e  s,d / a,e 

(A32)

Inside the deaerator, m s is in a liquid form. (ix) Boiler Feed Pump  w,i  m w, e m

(A33)

wp   vw,i . pe  pi 

(A34)

 w,i  w,i  m  w,e  w,e pump   w,i   pe  pi  X dest, p   m

(A35)

Note: Various mass, energy and exergy balance equations given in this section, for major components of the combined cycle system described in this paper, were modeled in the computational code developed in-house.

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List of Figures Fig. 1 Schematic of 3PR combined-cycle configuration with VA inlet air cooling (VA3PR) and without inlet air cooling (B3PR) Fig.2 (a) . Effect of ambient relative humidity and ambient temperature on plant specific work of 3PR configuration with and without VAIAC. Fig.2 (b) . Effect of ambient relative humidity and ambient temperature on plant efficiency of 3PR configuration with and without VAIAC. Fig.3. Variation of Exergy efficiency with TIT for 3PR cycle with and without inlet air cooling Fig 4. Exergy efficiency and Fuel Exergy Depletion Ratio of B3PR and VA3PR cycle. Fig 5 (a). Comparison of Fuel Exergy Depletion Ratio of 3PR cycle with and without vapor absorption inlet cooling at TIT = 1550 K. 44

Fig 5 (b). Comparison of Fuel Exergy Depletion Ratio of 3PR cycle with and without vapor absorption inlet cooling at TIT = 1750 K. Fig 6 (a). Fuel Exergy Depletion Ratio of VA3PR cycle for variation in TIT Fig 6 (b) . Fuel Exergy Depletion Ratio of VA3PR cycle for variation in βcomp Fig 7. Fuel Exergy Depletion Ratio of VA3PR cycle for variation in TIC Fig 8 (a) . Fuel Exergy Depletion Ratio of VA3PR cycle for variation in Tamb Fig 8 (b). Fuel Exergy Depletion Ratio of VA3PR cycle for variation in RHamb Fig 9. Exergetic improvement potential of major VA3PR combined cycle components. Fig 10. Relative Exergy Consumption Ratio and Productivity Lack Ratio of major VA3PR combined cycle Components. List of Tables Table 1. VA3PR Combined Cycle thermodynamic Data Table 2. Expressions of Exergy destruction rate for each component of system Table 3. Input data for analysis [8, 9]

45

Table 3. Input data for analysis [8, 9] PARAMETER SYMBOL Gas Properties: Enthalpy h=cp(T) dT Compressor i. Polytropic efficiency(pc) ii. Inlet plenum loss= Combustor

Gas turbine

HRSG

Values

i. Combustor efficiency (comb) ii. Pressure loss (ploss) iii. Lower heating value (LHV) i. Polytropic efficiency (pt) ii. Exhaust pressure iii. Exhaust hood loss iv. Turbine Blade Temperature

92.0 0.5% of entry pr.

UNIT kJ/kg % bar

99.5 2.0% of entry pressure 42.0 92.0 1.08 4 1123

% bar mJ/kg % bar K K

i. Effectiveness ii. Pressure loss=

98 10% of entry pressure (3PR (water side) parameters) iii. Pressure loss (gas side) 6% of entry pressure iv. Stack (minimum temperature) 353.0 v. H.P pressure 160 vi. H.P superheat 843(Max.) vii. H.P steam-turbine exhaust pressure 35 viii. Reheater loss 3% of entry pressure ix. I.P pressure 35.0 x. I.P superheat 603.0 xi. L.P pressure 6.0 xii. L.P superheat 573.0 xiii. Deaerator pressure 2.0 xiv. Condenser pressure 0.05 xv.Gas/steam approach temperature difference 20.0 i. Pinch-point temperature difference 10.0 Steam turbine i. Isentropic efficiency (HPT)= 88.0 ii. Isentropic efficiency (LPT)= 92 iii. Mechanical efficiency 98.5 iv. Minimum steam quality at LP exhaust 0.88 Inlet Air Cooling i. Refrigerant for VA system --System ii. Heat transfer effectiveness in VA generator 0.96 (εgen) iii. VA generator temperature 393 iv. Temp. of gas extracted from HRSG for generator of VAIAC system 423

%

Condenser Electric

K %

Undercooling Generator efficiency

4 98.5

Generator 46

bar bar K bar K bar bar bar K bar K bar bar K K % % % dry NH3 --K K

HIGHLIGHTS



A 3PR combined cycle with VAIAC is considered for analysis.



Component wise exergy destruction of IAC combined cycle is presented.



The influence of TIT, TIC , βcomp , Tamb , and RHamb on CC performance is assessed.



Inlet air cooling improves overall exergy efficiency of combined cycle system.

47