Articulo - Analisis de Energia y Exergia de Plantas de Potencia

E

Sa

b

a

ARA

KEERBC

C

1

soiec

o

1d

Renewable and Sustainable Energy Reviews 15 (2011) 1857–1872

Contents lists available at ScienceDirect

Renewable and Sustainable Energy Reviews

journa l homepage: www.e lsev ier .com/ locate / rser

nergy and exergy analyses of thermal power plants: A review

.C. Kaushika, V. Siva Reddya,∗, S.K. Tyagib

Centre for Energy Studies, Indian Institute of Technology Delhi, Hauz Khas, New Delhi 110016, IndiaSchool of Infrastructure Technology & Resource Management, Shri Mata Vaishno Devi University, Katra 182320 (J&K), India

r t i c l e i n f o

rticle history:eceived 2 October 2010

a b s t r a c t

The energy supply to demand narrowing down day by day around the world, the growing demand ofpower has made the power plants of scientific interest, but most of the power plants are designed by

ccepted 1 December 2010

eywords:xergynergyankine cyclerayton cycle

the energetic performance criteria based on first law of thermodynamics only. The real useful energyloss cannot be justified by the fist law of thermodynamics, because it does not differentiate betweenthe quality and quantity of energy. The present study deals with the comparison of energy and exergyanalyses of thermal power plants stimulated by coal and gas. This article provides a detailed review ofdifferent studies on thermal power plants over the years. This review would also throw light on the scopefor further research and recommendations for improvement in the existing thermal power plants.

o-generation © 2010 Elsevier Ltd. All rights reserved.

ontents

1. Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 18572. Energy and exergy analyses of coal fired power plants . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1858

2.1. Description of coal fired power plant . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 18582.2. Energy analysis . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 18582.3. Exergy analysis . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 18612.4. Results and discussion . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1865

3. Energy and exergy analysis of gas-fired combine cycle thermal power plants . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 18653.1. Description of gas-fired combine cycle thermal power plants . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 18653.2. Energy analysis . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 18653.3. Exergy analysis . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 18673.4. Results and discussion . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1869

4. Improvement potentials . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 18715. Conclusions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1871

References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 1872

. Introduction

Energy consumption is one of the most important indicatorhowing the development stages of countries and living standardsf communities. Population increment, urbanization, industrial-

fuels (coal, petroleum, fuel-oil, natural gas) fired thermal powerplants, whereas 20% of the electricity is compensated from differ-ent sources such as hydraulic, nuclear, wind, solar, geothermal andbiogas [1]. Generally, the performance of thermal power plants is

zing, and technologic development result directly in increasingnergy consumption. This rapid growing trend brings about therucial environmental problems

such as contamination and greenhouse effect. Currently, 80%f electricity in the world is approximately produced from fossil

∗ Corresponding author. Tel.: +91 9873592145.E-mail address: [email protected] (V.S. Reddy).

364-0321/$ – see front matter © 2010 Elsevier Ltd. All rights reserved.oi:10.1016/j.rser.2010.12.007

evaluated through energetic performance criteria based on firstlaw of thermodynamics, including electrical power and thermalefficiency. In recent decades, the exergetic performance based onthe second law of thermodynamics has found as useful methodin the design, evaluation, optimization and improvement of ther-mal power plants. The exergetic performance analysis can not
only determine magnitudes, location and causes of irreversibili-ties in the plants, but also provides more meaningful assessment ofplant individual components efficiency. These points of the exer-getic performance analysis are the basic differences from energetic
dx.doi.org/10.1016/j.rser.2010.12.007

http://www.sciencedirect.com/science/journal/13640321

http://www.elsevier.com/locate/rser

mailto:[email protected]

dx.doi.org/10.1016/j.rser.2010.12.007

1858 S.C. Kaushik et al. / Renewable and Sustainable

Nomenclature

m mass flow rate kg/sT temperature (◦C)W work done (kJ/s)Q heat transfer (kJ/s)E flow energy (kJ/s)ExQ exergetic heat (kJ/s)� flow exergy (kJ/s)S specific entropy (kJ/kg K)h Specific entropy (kJ/kg)Idestroyed irreversibility (kJ/s)

Subscriptsa airf fuelg gaso dead states steamw water

pedwttaiwa

meibicloduHoBsdpsi

ciSmCmemogf

I inletO outlet

erformance analysis. Therefore, it can be said that performingxergetic and energetic analyses together can give a completeepiction of system characteristics. Such a comprehensive analysisill be a more convenient approach for the performance evalua-

ion and determination of the steps towards improvement [4–6]. Inhe literature, there exist a number of papers concerning energeticnd exergetic performances of coal-fired thermal power plants. Fornstance, Hasan et al. [1] presented thermodynamic inefficiencies as

ell as reasonable comparison of each plant to others are identifiednd discussed for the coal-fired thermal power plants in Turkey.

Aljundi [2] determined the performance of the plant was esti-ated by a component wise modeling and a detailed break-up of

nergy and exergy losses for the considered steam power plantn Jordan. Datta et al. [3] presented exergy analysis of a coal-ased thermal power plant by splitting up the entire plant cycle

nto three zones for the analysis. Naterer et al. [4] analyzed theoal-fired thermal power plant with measured boiler and turbineosses. Rosen [5] presented energy and exergy-based comparisonf coal-fired and nuclear steam power plants. Ganapathy et al. [6]etermined the energy losses and the exergy losses of the individ-al components of the lignite fired thermal power plant. Zubair andabib [7] performed second law based thermodynamic analysisf the regenerative-reheat Rankine cycle power plants. Reddy andutcher [8] analyzed waste heat recovery based power generationystem based on second law of thermodynamics. Suresh et al. [9]etermined the exergetic performance of the coal-based thermalower plants using subcritical, supercritical, and ultra-supercriticalteam conditions. Oktay [10] presented exergy loss and proposedmproving methods for a fluidized bed power plant in Turkey.

Reddy and Mohamed [11] analyzed a natural gas fired combinedycle power generation unit to investigate the effect of gas turbinenlet temperature and pressure ratio on the exergetic efficiency.rinivas et al. [12] analyzed the combined cycle power plant usingethane as a fuel using the first and second law of thermodynamics.

an and Smith [13] described a simple but accurate method to esti-ate the Rankine bottoming cycle power output directly from the

xergy of the gas turbine exhaust, utilizing the second law of ther-odynamics. Datta et al. [14] presented energy and exergy analyses

f an externally fired gas turbine cycle integrated with biomassasifier for distributed power generation. Sue and Chuang [15] per-ormed the engineering design and theoretical exergetic analysis of

Energy Reviews 15 (2011) 1857–1872

the combustion gas turbine based power generation systems. Bil-gen [16] presented the exergetic and engineering analyses as wellas a simulation of gas turbine-based cogeneration plants consistingof a gas turbine, heat recovery steam generator and steam turbine.Khaliq and Kaushik [17] presented the second-law approach for thethermodynamic analysis of the reheat combined Brayton/Rankinepower cycle. Woudstra et al. [18] determined the cogeneration pro-cess, levels of steam generation to reduce the heat transfer lossesin the heat recovery steam generator (HRSG) and the exergy lossdue to the exhaust of flue gas to the stack.

Keeping in view the facts stated above, it can be expected thatperforming an analysis based on the same definition of perfor-mance criteria will be meaningful for performance comparisons,assessments and improvement for thermal power plants. Addi-tionally, considering both the energetic and exergetic performancecriteria together can guide the ways of efficient and effective usageof fuel resources by taking into account the quality and quantityof the energy used in the generation of electric power in thermalpower plants. The purpose of this study presented here is to carryout energetic and exergetic performance analyses, at the designconditions, for the existing coal and gas-fired thermal power plantsin order to identify the needed improvement. For performing thisaim, we summarized thermodynamic models for the consideredpower plants on the basis of mass, energy and exergy balance equa-tions. The thermodynamic model simulation results are compared.In the direction of the comprehensive analysis results, the require-ments for performance improvement are evaluated.

2. Energy and exergy analyses of coal fired power plants

Coal based thermal power plant generally operates on Rankinecycle. In ideal vapor power cycle many, such as Carnot cycle hasimpracticalities associated with it can be eliminated by superheat-ing the steam in the boiler and condensing it completely in thecondenser [19].

2.1. Description of coal fired power plant

We are considering the analysis of a cumulative coal fired ther-mal power plant with all methods of the efficiency increasingtechnics like lowering the condenser pressure, superheating thesteam to high temperatures, increasing the boiler pressure, reheatand regenerative Rankine cycle, as shown in Fig. 1. A continuousmass flow diagram for one unit of the power plant modeled in thisstudy includes the main components such as high, intermediateand low pressure turbines (HPT, IPT and LPT), a boiler (B), number ofpumps (P), a dearetor (D), a generator (G), a condenser (C), low andhigh pressure feed water heaters (LPH and HPH). The thermody-namic models are based on fundamental mass and energy balances.Using the energy and mass balance equations for each componentin the power plant model, it is possible to compute energy andexergy contents in terms of turbine power outputs, pump powerconsumptions, boiler heat requirements, energy and exergy flowsat each node of the plants, component first and second efficiencies,component irreversibilities in the plants, and so on.

2.2. Energy analysis

In an open flow system there are three types of energy transferacross the control surface namely working transfer, heat transfer,and energy associated with mass transfer and/or flow. The first lawof thermodynamics or energy balance for the steady flow process

S.C. Kaushik et al. / Renewable and Sustainable Energy Reviews 15 (2011) 1857–1872 1859

therm

o

∑

waCs

ci

�

Tac

Fig. 1. Coal-fired

f an open system is given by:

Qk + m

(hi + C2

i

2+ gZi

)= m

(ho + C2

o

2+ gZo

)+ W

here Qk heat transfer to system from source at temperature Tk,nd W is the net work developed by the system. The other notationsis the bulk velocity of the working fluid, Z, is the altitude of the

tream above the sea level, g is the specific gravitational force.The energy or first law efficiency �I of a system and/or system

omponent is defined as the ratio of energy output to the energynput to system/component i.e.

I = Desired output energyInput energy supplied

o analyze the possible realistic performance, a detailed energynalysis of the coal fired thermal power plant system has beenarried out by ignoring the kinetic and potential energy change

(a) The energy balance for boiler:The energy balance for the combustion/boiler is give by:

0 = Qk − mw(h10 − h9) − ms(h15 − h14)] − Energy loss

where mw is the mass flow rate of water, ms is the mass flowrate of steam combustion which gives:

Energy loss = Qk − mw(h10 − h9) − ms(h15 − h14)

al power plant.

The first law efficiency is defined as

�I,Boiler = Energy outputEnergy input

= 1 − Energy lossEnergy input

= mw(h10 − h9) + ms(h15 − h14)

Qk

(b) The energy balance for the high pressure turbine is give by:

WHPT = m10(h10 − h11) + (m10 − m11)(h11 − h14) − Energy loss

This gives:

Energy loss = m10(h10 − h11) + (m10 − m11)(h11 − h14) − WHPT

The first law efficiency is:

�I,HPT = 1 − Energy lossm10(h10 − h11) + (m10 − m11)(h11 − h14)

= WHPT

m10(h10 − h11) + (m10 − m11)(h11 − h14)

(c) The energy balance for the intermediate pressure turbine isgive by:

WIPT = m15(h15 − h20) + (m15 − m20)(h20 − h23)

+ (m − m − m )(h − h ) − Energy loss
15 20 23 23 19
This gives:

Energy loss = m15(h15 − h20) + (m15 − m20)(h20 − h23)

+ (m15 − m20 − m23)(h23 − h19) − WIPT

1 inable

m20

− m

h27)

860 S.C. Kaushik et al. / Renewable and Susta


�I,IPT = 1 − Energy lossm15(h15 − h20) + (m15 − m20)(h20 − h23) + (m15 −

= WIPT

m15(h15 − h20) + (m15 − m20)(h20 − h23) + (m15 − m20

(d) The energy balance for the low pressure turbine is give by:

WLPT = m19(h19 − h27)

+ (m19 − m27)(h27 − h30) − Energy loss

This gives:

Energy loss = m19(h19 − h27) + (m19 − m27)(h27 − h30) − WLPT


�I,LPT = Energy lossm19(h19 − h27) + (m19 − m27)(h27 − h30)

=m19(h19 −

Condenser sub system(e) The energy balance for the condenser is give by:

0 = m30(h30 − h1) − Qk − Energy loss

This gives:

Energy loss = m30(h30 − h1) − Qk


�I,Cond = 1 − Energy lossm30(h30 − h1)

Pump sub system(f) The energy balance for the low pressure pump is give by:

−WLPP = m1(h1 − h2) − Energy loss

This gives:

Energy loss = m1(h1 − h2) + WLPP


�I,LPP = 1 − Energy lossWLPP

= m1(h2 − h1)WLPP

(g) The energy balance for the high pressure pump is give by:

−WHPP = m5(h5 − h6) − Energy loss

This gives:

Energy loss = m5(h5 − h6) + WHPP


�I,HPP = 1 − Energy lossWHPP

= m1(h6 − h5)WHPP

Energy Reviews 15 (2011) 1857–1872

− m23)(h23 − h19)

23)(h23 − h19)

WLPT

+ (m19 − m27)(h27 − h30)

Feed water heater sub system(h) The energy flow equation for the high presure feed water

heater (HPH1) system becomes:

0 = m11(h11 − h12) − m8(h9 − h8) − Energy loss

This gives:

Energy loss = [m11(h11 − h12) − m8(h9 − h8)]


�I,HPH1 = 1 − Energy lossm11(h11 − h12)

= m8(h9 − h8)m11(h11 − h12)

(i) The energy flow equation for the high presure feed waterheater (HPH2) system becomes:


This gives:




= m7(h8 − h7)m16(h16 − h17)

(j) The energy flow equation for the high presure feed waterheater (HPH3) system becomes:


This gives:




= m6(h7 − h6)m20(h20 − h21)

(k) The energy flow equation for the low presure feed water heater(LPH1) system becomes:

˙ ˙
This gives:


inable

(

2

toestaettig

W

EpaaEimq

S.C. Kaushik et al. / Renewable and Susta


�I,LPH1 = 1 − Energy lossm24(h24 − h25)

= m3(h4 − h3)m24(h24 − h25)

(l) The energy flow equation for the low presure feed water heater(LPH2) system becomes:


This gives:



�II,LPH2 = 1 − Energy lossm27(h27 − h28)

= m2(h3 − h2)m27(h27 − h28)

m) Dearetor sub systemIt is an adiabatic mixing chamber where a hot streams 23,

22 are mixed with a cold stream 4, forming a mixture 5, theenergy supplied is the sum of the energies of the hot and coldstreams, and the energy recovered is the energy of the mixture.The energy flow equation for the dearetor system becomes:

0 = m22h22 + m23h23 + m4h4 − m5h5 − Energy loss

where m5 = m22 + m23 + m24. This gives:

Energy loss = m22h22 + m23h23 + m4h4 − m5h5


�II,Der = 1 − Energy lossm22h22 + m23h23 + m4h4

= m5h5

m22h22 + m23h23 + m4h4

.3. Exergy analysis

Exergy is a generic term for a group of concepts that definehe maximum possible work potential of a system, a streamf matter and/or heat interaction; the state of the (conceptual)nvironment being used as the datum state. In an open flowystem there are three types of energy transfer across the con-rol surface namely working transfer, heat transfer, and energyssociated with mass transfer and/or flow. The work transfer isquivalent to the maximum work, which can be obtained fromhat form of energy. The exergy (� Q) of heat transfer Q fromhe control surface at temperature T is determined from max-mum rate of conversion of thermal energy to work Wmax. isiven by:

max = �Q = Q(

1 − T0

T

)

xergy of steady flow stream of matter is the sum of kinetic,otential and physical exergy. The kinetic and potential energyre almost equivalent to exergy. The physical specific exergy � i
nd � o depends on initial state of matter and environmental state.nergy analysis is based on the first law of thermodynamics, whichs related to the conservation of energy. Second law analysis is a
ethod that uses the conservation of mass and degradation of theuality of energy along with the entropy generation in the analysis

Energy Reviews 15 (2011) 1857–1872 1861

design and improvement of energy systems. Exergy analysis is auseful method; to complement but not to replace energy analysis.

The exergy flow for steady flow process of an open system isgiven by

∑(1 − T0

Tk

)Qk +

∑in

m�i = �W +∑out

m�o + Idestroyed;

�m[(h0 − h0o) − T0(s − s0)], h0

= h + C2

2+ gZ, Idestroyed = T0[Sgen]

where � i and � o are exergy associated with mass inflow andoutflows are respectively, � W is useful work done on/by system,Idestroyed is irreversibility of process and h0 is the methalpy as sum-mation of enthalpy, kinetic energy and potential energy. The othernotations C is the bulk velocity of the working fluid, Z is the alti-tude of the stream above the sea level, g is the specific gravitationalforce. The irreversibility may be due to heat transfer through finitetemperature difference, mixing of fluids at different temperatureand mechanical friction. Exergy analysis is an effective means, topinpoint losses due to irreversibility in a real situation.

The second law efficiency is defined as

�II = Actual thermal efficiencymaximum possible (reversible) thermal efficiency

= Exergy outputExergy input

To analyze the possible realistic performance, a detailed exergyanalysis of the coal fired thermal power plant has been carried outby ignoring the kinetic and potential energy change. For steadystate flow the exergy balance for a thermal system is given as below[20]:

�W =n∑

k=1

(1 − T0

Tk

)Qk +

r∑k=1

[(m� )i − (m� )o]k − T0Sgen

where � W represents the useful work done and/or by the system,the first term on the right hand side [(1 − T0/Tk)Qk] represents theexergy summation supplied through heat transfer, while changesin the exergy summation of the working fluid is represented bythe second term

∑[(m� )i − (m� )o] where i and o refers the inlet

and outlet states. On the other hand, the exergy distraction and/orthe irreversibility in the system is given by the last term on theright hand side, [T0Sgen]. The other notations such as, Q is the heattransfer rate, m is the mass flow rate of the working fluid, � isthe exergy flow rate per unit mass, Sgen is the entropy generationrate, T0 is the ambient air temperature, TK is the temperature ofthe heat source/sink at which the heat is transferred/rejected. Thecomponent wise exergy balance of the coal fire thermal power plantsystem is given as below.

(a) The exergy balance for boiler combustion:The exergy balance for the combustion/boiler is give by:

0 =r∑

k=1

[(m� )f +a − (m� )p]k

− T0Sgen

where mf+a is sum of the mass of coal and air, mp is products aftercombustion which gives:

T0Sgen = [(m� )f +a − (m� )i]

Entropy of the flue gas and hot products are obtained using TableA-18 to Table A-20 and Table A-27 from Cengel and Michael [19].

1862 S.C. Kaushik et al. / Renewable and Sustainable Energy Reviews 15 (2011) 1857–1872

ructio

�

e

0

T

I

�

�

Ft

Fig. 2. Comparison of heat losses and exergy dest


III = Exergy outputExergy input

= 1 − Exergy lossExergy input

= 1 − T0Sgen

(m� )f +a= (m� )p

(m� )f +a

(b) The exergy balance for high temperature heat exchangerThe exergy flow equation for the High temperature heat

xchanger becomes:

= mp(�i − �O) − mw(�10 − �9) − ms(�15 − �14) − T0Sgen

This gives:

0Sgen = [{mp(hi − h0) − mw(h10 − h9) − ms(h15 − h14)}−T0{mp(si − s0) − mw(s10 − s9) − ms(s15 − s14)}]

The irreversibility = exergy loss is

˙destroyed = T0Sgen

The second law efficiency is:

II,heat = 1 − Idestroyed = mw(�10 − �9) + ms(�15 − �14)
mp(�i − �o) mp(�i − �o)
Total boiler subsystem second law efficiencies is

II,Boiler = mw(�10 − �9) + ms(�15 − �14)(m� )f

ig. 3. Values of thermodynamic loss rate to capital cost ratio, R, for several devices in ao 2002 US dollars (as explained in the text). Note that * shows R = 85:8 W/$ based on ene

n (kW) in major components of the power plant.

(c) The exergy balance for the high pressure turbine is give by:

WHPT = m10(�10 − �11) + (m10 − m11)(�11 − �14) − T0Sgen

This gives:

T0Sgen = m10(�10 − �11) + (m10 − m11)(�11 − �14) − WHPT

and the entropy generation rate is:

Sgen = m10(s11 − s10) + (m10 − m11)(s14 − s11)

The irreversibility = exergy loss is:

Idestroyed = T0Sgen = T0[m10(s11 − s10) + (m10 − m11)(s14 − s11)]


�II,HPT = 1 − Idestroyed

m10(�10 − �11) + (m10 − m11)(�11 − �14)

= WHpT

m10(�10 − �11) + (m10 − m11)(�11 − �14)

(d) The exergy balance for the Intermediate pressure turbine isgive by:

WIPT = m15(�15 − �20) + (m15 − m20)(�20 − �23)

+ (m15 − m20 − m23)(�23 − �19) − T0Sgen

This gives:

T0Sgen = m15(�15 − �20) + (m15 − m20)(�20 − �23)

500 MW unit of a coal fired electrical generating station. Costs have been modifiedrgy loss for the condenser.

inable

a

S

I

�20 −

m23

W

T

a

S

I

�

0

T

I

�

−


+ (m15 − m20 − m23)(�23 − �19) − WIPT

nd the entropy generation rate is:

˙ gen = m15(s20 − s15) + (m15 − m20)(s23 − s20)

+ (m15 − m20 − m23)(s19 − s23)


˙destroyed = T0Sgen = T0[m15(s20 − s15) + (m15 − m20)(s23 − s20)

+ (m15 − m20 − m23)(s19 − s23)]


II,IPT = 1 − Idestroyed

m15(�15 − �20) + (m15 − m20)(�20 − �23) + (m15 − m

= WIpT

m15(�15 − �20) + (m15 − m20)(�20 − �23) + (m15 − m20 −

(e) The exergy balance for the Low pressure turbine is give by:

LPT = m19(�19 − �27) + (m19 − m27)(�27 − �30) − T0Sgen

This gives:

0Sgen = m19(�19 − �27) + (m19 − m27)(�27 − �30) − WLPT


˙ gen = m19(s27 − s19) + (m19 − m27)(s30 − s27)


˙destroyed = T0Sgen = T0[m19(s27 − s19) + (m19 − m27)(s30 − s27)]


II,LPT = 1 − Idestroyed

m19(�19 − �27) + (m19 − m27)(�27 − �30)

= WLpT

m19(�19 − �27) + (m19 − m27)(�27 − �30)

Condenser sub system(f) The exergy balance for the condenser is give by:

= m30(�30 − �1) −n∑

k=1

(1 − T0

Tk

)Qk − T0Sgen

This gives:

0Sgen = m30(�30 − �1) −n∑

k=1

(1 − T0

Tk

)Qk


˙destroyedT0Sgen

= [{m30(h30 − h1)} − T0{m30(s30 − s1)}] −n∑

k=1

(1 − T0

Tk

)Qk


II,Con = 1 − Idestroyed

m30(�30 − �1)

Pump sub system(g) The exergy balance for the low pressure pump is give by:

WLPP = m1(�1 − �2) − T0Sgen

Energy Reviews 15 (2011) 1857–1872 1863

m23)(�23 − �19)

)(�23 − �19)

This givesThe irreversibility = exergy loss is:

Idestroyed = T0Sgen = m1(�1 − �2) + WLPP = m1T0(s2 − s1)


�II,LPP = 1 − Idestroyed

WLPP= m1(�2 − �1)

WLPP

(h) The exergy balance for the High pressure pump is give by:

−WHPP = m5(�5 − �6) − T0Sgen

This gives:The irreversibility = exergy loss is:

Idestroyed = T0Sgen = m5(�5 − �6) + WHPP = m5T0(s6 − s5)


�II,HPP = 1 − Idestroyed

WHPP= m1(�6 − �5)

WHPP

Feed water heater sub system(i) The exergy flow equation for the high presure feed water

heater (HPH1) system becomes [29]:

0 = m11(�11 − �12) − m8(�9 − �8) − T0Sgen

This gives:

T0Sgen = [{m11(h11 − h12) − m8(h9 − h8)}− T0{m11(s11 − s12) − m8(s9 − s8)}]


Idestroyed = T0Sgen


�II,HPH1 = 1 − Idestroyed

m11(�11 − �12)= m8(�9 − �8)

m11(�11 − �12)

(j) The exergy flow equation for the high presure feed waterheater (HPH2) system becomes:

0 = m16(�16 − �17) − m7(�8 − �7) − T0Sgen

This gives:

T0Sgen = [{m16(h16 − h17) − m7(h8 − h7)}− T0{m16(s16 − s17) − m7(s8 − s7)}]


Idestroyed = T0Sgen


�II,HPH2 = 1 − Idestroyed = m7(�8 − �7)
m16(�16 − �17) m16(�16 − �17)
(k) The exergy flow equation for the high presure feed waterheater (HPH3) system becomes:

0 = m20(�20 − �21) − m6(�7 − �6) − T0Sgen

1 inable

T

I

�

h

0

T

I

�

h

0

T

I

�

aste

0

w

T

a

S

I


This gives:

0Sgen = [{m20(h20 − h21) − m6(h7 − h6)}− T0{m20(s20 − s21) − m6(s7 − s6)}]




II,HPH3 = 1 − Idestroyed

m20(�20 − �21)= m6(�7 − �6)

m20(�20 − �21)

(l) The exergy flow equation for the low presure feed watereater (LPH1) system becomes:

= m24(�24 − �25) − m3(�4 − �3) − T0Sgen

This gives:

0Sgen = [{m24(h24 − h25) − m3(h4 − h3)} − T0{m24(s24 − s25)

− m3(s4 − s3)}]




II,LPH1 = 1 − Idestroyed

m24(�24 − �25)= m3(�4 − �3)

m24(�24 − �25)

(m) The exergy flow equation for the low presure feed watereater (LPH2) system becomes:

= m27(�27 − �28) − m2(�3 − �2) − T0Sgen

This gives:





II,LPH2 = 1 − Idestroyed

m27(�27 − �28)= m2(�3 − �2)

m27(�27 − �28)

(n) Dearetor sub systemIt is an adiabatic mixing chamber where a hot streams 23, 22

re mixed with a cold stream 4, forming a mixture 5, the exergyupplied is the sum of the exergies of the hot and cold streams, andhe exergy recovered is the exergy of the mixture. The exergy flowquation for the dearetor system becomes [30]:

= m22�22 + m23�23 + m4�4 − m5�5 − T0Sgen

here m5 = m22 + m23 + m4This gives:

0Sgen = m22�22 + m23�23 + m4�4 − m5�5


˙ gen = m5s5 − m22s22 − m23s23 − m4s4


˙destroyed = T0Sgen = T0[m5s5 − m22s22 − m23s23 − m4s4]

Energy Reviews 15 (2011) 1857–1872


Idestroyed = T0Sgen


�II,Der = 1 − T0Sgen

m22�22 + m23�23 + m4�4

= m5�5

m22�22 + m23�23 + m4�4

Expansion valve(o) The exergy flow equation for the expansion valve (EXP1)

becomes:

0 = m12(�12 − �13) − T0Sgen

This gives:

T0Sgen = m12[(h12 − h13) − T0(s12 − s13)]

The irreversibility = exergy loss:

Idestroyed = T0Sgen


�II,EXP1 = 1 − T0Sgen

m12(�12 − �13)

(p) The exergy flow equation for the expansion valve (EXP2)becomes:

0 = m17(�17 − �18) − T0Sgen

This gives:

T0Sgen = m17[(h17 − h18) − T0(s17 − s18)]


Idestroyed = T0Sgen



m17(�17 − �18)

(q) The exergy flow equation for the expansion valve (EXP3)becomes:

0 = m21(�21 − �22) − T0Sgen

This gives:

T0Sgen = m21[(h21 − h22) − T0(s21 − s22)]


Idestroyed = T0Sgen



m21(�21 − �22)

(r) The exergy flow equation for the expansion valve (EXP4)becomes:

0 = m25 − (�25 − �26) − T0Sgen

This gives:

T0Sgen = m25[(h25 − h26) − T0(s25 − s26)]


Idestroyed = T0Sgen

inable

�

b

0

T

I

�

2

ttfxeovdsTa[lNiFcbifplt

edupfrtda7aadiFi

oi



II,EXP4 = 1 − T0Sgen

m25(�25 − �26)

(s) The exergy flow equation for the expansion valve (EXP5)ecomes:

= m28(�28 − �29) − T0Sgen

This gives:

0Sgen = m28[h28 − h29) − T0(s28 − s29)]




II = 1 − T0Sgen

m28(�28 − �29)

.4. Results and discussion

The flow availabilities of different species at inlet and outlet ofhe Boiler have been evaluated with respect to an exergy referencehermodynamic state of Pr = 101.35 kN/m2, Tr = 298.15 K with moleractions of the constituents as xO2

r = 0.2035, xCO2r = 0.0003, and

H2Or = 0.0303 as recommended by Moran and Shapiro [21]. Somt al. [22] prepared a theoretical model of exergy balance, basedn availability transfer and flow availability, in the process of pul-erized coal combustion in a tubular air-coal combustor has beeneveloped to evaluate the total thermodynamic irreversibility andecond law efficiency of the process at various operating conditions.he fuel considered in the present analysis is coal, whose ultimatenalysis is as follows: 70.2% C, 5.7% H, 13.4% O, 1.9% N, and 8.8% ash23]. He noticed that as the inlet air pressure increases the secondaw efficiency decreases but the combustion efficiency is increases.aterer et al. [4] presented energy and exergy analysis of subcrit-

cal boiler–turbine generator for a 32 MW coal-fired power plant.rom Fig. 2, it can be observed that although the energy loss in theondenser seems higher but, the largest exergy losses occur in theoiler with the highest exergy destruction in the existing plant. This

llustrates the importance of exergy analysis, as it provides the dif-erent insight and trends than that of the energy analysis. Someoneerforming an energy analysis would led to believe that the largest

osses occur in the condenser, whereas the exergy analysis proveshat they occur in the boiler.

Dincer and Rosen [24] demonstrated that, although energy andxergy values are dependent on the intensive properties of theead state, the main results of energy and exergy analyses aresually not significantly sensitive to reasonable variations in theseroperties [25]. In some extreme cases, such as a rocket taking offrom the ground level and flying to space, the evaluation of accu-ate energy and exergy values requires to be taken of care becausehe variations in dead-state properties are large. Saidur et al. [26]etermined energy and exergy efficiencies have been determineds well. In a boiler, the energy and exergy efficiencies are found to be2.46% and 24.89%, respectively. Dincer and Rosen [27] presentedsystematic correlation appears to exist between exergy loss ratend capital cost for that purpose they have taken a thermodynamicata for the coal fired Nanticoke Generating Station consists of eight

ndividual units, each having approximatly net outputs of 500 MW.
ig. 3 shows the exergy and energy loss rate to capital cost ratio ofndividual components in the power plant.
It may be mentioned that so far only few studies have been donen exergy and energy analysis in coal fired thermal power plant. Its observed that in most of the cases, the major portion of exergy

Energy Reviews 15 (2011) 1857–1872 1865

is lost in the combustor of a boiler. So, it should be taken into con-siderations for minimizing the losses in the combustion chamber.It may be due to incomplete combustion, improper insulation andentropy generation in this sub system. Table 1 shows the compari-son of energy and exergy efficiencies and losses in coal fired thermalpower plant with other works available in the literature.

3. Energy and exergy analysis of gas-fired combine cyclethermal power plants

Gas-fired combine cycle thermal power plant toping cycle basedon the Brayton cycle and bottoming cycle based on the Rankinecycle.

3.1. Description of gas-fired combine cycle thermal power plants

The Brayton cycle was first proposed by George Brayton in 1870for the use in the reciprocating oil-burning engine. Today, it is usedfor gas turbines only where both the compression and expansionprocesses take place in rotating machinery [28]. The total reversibleprocesses cycle shown schematically on a T–s diagram in Fig. 4. Theearly gas turbines built in the 1940s and even 1950s had simple-cycle efficiencies of about 17% because of the low compressor andturbine efficiencies and low turbine inlet temperatures due to met-allurgical limitations of those times [31]. Increasing the turbineinlet (or firing) temperatures this has been the primary approachtaken to improve gas-turbine efficiency. The turbine inlet tem-peratures have increased steadily from about 540 ◦C (1000 ◦F) inthe 1940s to 1425 ◦C (2600 ◦F) and even higher today [19]. Weare considering for analysis cumulative gas fired combined cyclethermal power plant with all methods of the efficiency enhance-ment like increasing turbine inlet temperature and pressure, multypressure heat recovery steam generator, increasing the boiler pres-sure, reheat and regenerative Rankine cycle, as shown in Fig. 5. Thecontinuous mass flow diagram for one of the units of any powerplant modeled in this study includes the main components such as,gas turbine (T), high, intermediate and low pressure turbine (HPT,IPT and LPT), heat recovery steam generator (HRSG), several feedwater pumps (P), a dearetor, a generator (G), a condenser (COND),feed water heater, compressor (C), combustion chamber (CC). Usingthe balance energy and mass equations for each component in thepower plant, energy, exergy flows and at each node of the plant canbe calculated the numerically as well as analytically, for given setof operating conditions.

3.2. Energy analysis

To analyze the possible realistic performance, a detailed energyanalysis of the gas fired combined cycle thermal power plant hasbeen carried out by ignoring the kinetic and potential energychange. For steady state flow the energy balance for a thermalsystem is given as below:

∑Qk + m

(hi + C2

i

2+ gZi

)= m

(ho + C2

o

2+ gZo

)+ W

where Qk heat transfer to system from source at temperature Tk,and W is the net work developed by the system. The other notationsC is the bulk velocity of the working fluid, Z, is the altitude of thestream above the sea level, g is the specific gravitational force.

The energy or first law efficiency �I of a system and/or system
component is defined as the ratio of energy output to the energyinput to system/ component i.e.
�I = Desired output energyInput energy supplied


Table 1Comparison of energy and exergy efficiencies.

S. no Power plants Capacity (MW) Exergy (second law) efficiency �I (%) Energy (first law) efficiency �II (%)Reference

Boiler Turbine Overall Boiler Turbine Overall

1 Yatagan 630 40.84 80.1 31.95 83.98 37.012 Seyitomer 600 36.75 85.45 37.88 94.44 38.033 Can 320 48.23 90.03 28.55 89.57 42.124 Catalagzi 300 45.47 88.6 35.49 88.25 37.88 [1–9]5 Kangal 457 36.45 84.19 33.09 92.10 37.19 Hasan et al. [1]6 Afsin Elbistan 1440 39.00 86.16 25.12 93.29 42.647 Orhaneli 210 45.77 64.42 26.95 85.51 37.638 Soma 990 41.43 86 37.36 90.21 36.089 Tuncbilek 429 44.00 86.36 31.50 89.57 38.4410 Zarqa 396 43.8 73.5 35.19 93.76 26.34 Aljundi [2]11 Neyveli 50 58.62 81.2 32.46 91.9 26.91 Ganapathy et al. [6]12 NGS 500 49.51 – 32.35 94.67 35.81 Dincer and Rosen [24]

turbi

−

E

�

Fig. 4. Gas

(a) The energy balance for air compressor sub systemThe energy balance for the air compressor is give by:

WC = ma(ha − hb) − Energy loss

This gives:

nergy loss = −ma(ha − hb) + WC


I,C = 1 − Energy lossWC

= ma(hb − ha)WC

Fig. 5. Gas-fired combined cyc

ne engine.

(b) The energy balance for combustion chamber sub systemThe energy balance for the combustion chamber is give by:

0 =r∑

k=1

[(mh)f +a) − (mh)p)]k − Energy loss

where mf+a is sum of the mass of gas and air, mp is products aftercombustion.

which gives:

Energy loss = [(mh)f +a) − (mh)i)]

le thermal power plant.

inable

�

W

E

�

(

b

0

E

�

W

E

�

W

E

�


The first law efficiency is

I,CC = Energy outputEnergy input

= 1 − Energy lossEnergy input

= 1 − Energy loss(mh)f +a

= (mh)p

(mh)f +a

(c) The energy balance for gas turbine sub systemThe energy balance for the gas turbine is give by:

GT = mp(hd − hc) − Energy loss

This gives:

nergy loss = mp(hd − hc) − WGT


I,GT = 1 − Energy lossmp(hd − hc)

= WGT

mp(hd − hc)

(d) The energy balance for heat recovery steam generatorHRSG) sub system

The energy flow equation for the boiler heat exchangerecomes:

= mp(hi − ho) − m4(h4 − h3) − m8(h10 − h8)

− m6(h7 − h6) − Energy loss

This gives:

nergy loss = [mp(hi − ho) − m4(h4 − h3)

− m8(h10 − h8) − m6(h7 − h6)]


I,HRSG = 1 − Energy lossmp(hi − ho)

= m4(h4 − h3) + m8(h10 − h8) + m6(h7 − h6)mp(hi − ho)

Steam turbine sub system(e) The energy balance for the high pressure turbine is give by:

HPT = m7(h8 − h7) − Energy loss

This gives:

nergy loss = m7(h8 − h7) − WHPT


I,HPT = 1 − Energy lossm7(h8 − h7)

= WHPT

m7(h8 − h7)

(f) The energy balance for the low pressure turbine is give by:

LPT = m10(h11 − h10) + (m10 − m11)(h12 − h11) − Energy loss

This gives:

nergy loss = m10(h11 − h10) + (m10 − m11)(h12 − h11) − WLPT


I,LPT = 1 − Energy lossm (h − h ) + (m − m )(h − h )
10 11 10 10 11 12 11
= WLPT

m10(h11 − h10) + (m10 − m11)(h12 − h11)

(g) Condenser sub system

Energy Reviews 15 (2011) 1857–1872 1867

The energy balance for the condenser is give by:

0 = m12(h12 − h1) − Qk − Energy loss

This gives:

Energy loss = m12(h12 − h1) − Qk


�I,Con = 1 − Energy lossm12(h12 − h1)

Pump sub system(h) The energy balance for the low pressure pump is give by:


This gives:



�I,LPP = 1 − Energy lossWLPP

= m1(h2 − h1)WLPP

(i) The energy balance for the High pressure pump is give by:


This gives:



�I,HPP = 1 − Energy lossWHPP

= m5(h6 − h5)WHPP

Feed water heater sub system(j) The energy flow equation for the low presure feed water

heater (LPH) system becomes:


This gives:



�I,LPH = 1 − Energy lossm11(h11 − h13)

= m2(h3 − h2)m11(h11 − h13)

(k) Dearetor sub systemThe energy flow equation for the dearetor system becomes:

0 = m9h9 + m4h4 − m5h5 − Energy loss

where m5 = m9 + m4. This gives:

Energy loss = m9h9 + m4h4 − m5h5


�I,Dearetor = 1 − Energy lossm9h9 + m4h4

= m5h5

m9h9 + m4h4

3.3. Exergy analysis

To analyze the possible realistic performance, a detailed exergyanalysis of the gas fired combined cycle thermal power plant hasbeen carried out by ignoring the kinetic and potential energy

1 inable

cs

�

wteitatrteTsnp

−

T

a

S

I

�

0

wc

T

A

�

W

T

a

S


hange. For steady state flow the exergy balance for a thermalystem is given as below:

W =n∑

k=1

(1 − T0

Tk

)Qk +

r∑k=1

[(m� )i − (m� )o]k − T0Sgen

here � W represents the useful work done and/or by the system,he first term on the right hand side [(1 − (T0/Tk))Qk] represents thexergy summation supplied through heat transfer, while changesn the exergy summation of the working fluid is represented byhe second term

∑[(m� )i − (m� )o] where i and o refer the inlet

nd outlet states. On the other hand, the exergy distraction and/orhe irreversibility in the system is given by the last term on theight hand side, [T0Sgen] The other notations suchas, Q is the heatransfer rate, m is the mass flow rate of the working fluid, � is thexergy flow rate per unit mass, Sgen is the entropy generation rate,0 is the ambient air temperature, TK is the temperature of the heatource/sink at which the heat is transferred/rejected. The compo-ent wise exergy balance of the gas fire combined cycle thermalower plant system is given as below.

(a) The exergy balance for air compressor sub systemThe exergy balance for the air compressor is give by:

WC = ma(�a − �b) − T0Sgen

This gives:

0Sgen = ma(�a − �b) + WC


˙ gen = ma(Sb − Sa)


˙destroyed = T0Sgen = T0[ma(Sb − Sa)]


II,C = 1 − Idestroyed

WC= ma(�b − �a)

WC

(b) The exergy balance for combustion chamber sub system

=r∑

k=1

[(m� )f +a − (m� )p]k

− T0Sgen

here mf+a is sum of the mass of gas and air, mp is products afterombustion which gives:

0Sgen = [(m� )f +a − (m� )i]

Entropy of the flue gas and hot products are obtained using Table-18 to Table A-20 and Table A-27 from Cengel and Michael [19].


II,CC = Exergy outputExergy input

= 1 − Exergy lossExergy input

= 1 − T0Sgen

(m� )f +a= (m� )p

(m� )f +a

(c) The exergy balance for gas turbine sub systemThe exergy balance for the gas turbine is give by:

GT = mp(�d − �c) − T0Sgen

This gives:

0Sgen = mp(�d − �c) − WGT


˙ gen = mp(Sc − Sd)

Energy Reviews 15 (2011) 1857–1872


Idestroyed = T0Sgen = T0[mp(Sc − Sd)]


�II,GT = 1 − Idestroyed

mp(�d − �c)= WGT

mp(�d − �c)

(d) The energy balance for heat recovery steam generator(HRSG) sub system

The performance of the HRSG strongly affects the overall per-formance of the combined cycle power plant. HRSG is nothing butshell and tube heat exchanger, in that hot gas flow through the shelland water flow thorough tubes.

The exergy flow equation for the boiler heat exchangerbecomes:

0 = mp(�i − �o) − m4(�4 − �3) − m8(�10 − �8)

− m6(�7 − �6) − T0Sgen

This gives:

T0Sgen = [{mp(hi − ho) − m4(h4 − h3) − m8(h10 − h8)

− m6(h7 − h6)} − T0{mp(si − so) − m4(s4 − s3)

− m8(s10 − s8) − m6(s7 − s6)}]


Idestroyed = T0Sgen


�II,HRSG = 1 − Idestroyed

mp(�i − �o)

= m4(�4 − �3) + m8(�10 − �8) + m6(�7 − �6)mp(�i − �o)

Steam turbine sub system(e) The exergy balance for the High pressure turbine is give by:

WHPT = m7(�8 − �7) − T0Sgen

This gives:

T0Sgen − m7(�8 − �7) − WHPT


Sgen = m7(s7 − s8)


Idestroyed = T0Sgen = T0[m7(s7 − S8)]


�II,HPT = 1 − Idestroyed

m7(�8 − �7)= WHPT

m7(�8 − �7)

(f) The exergy balance for the Low pressure turbine is give by:

WLPT = m10(�11 − �10) + (m10 − m11)(�12 − �11) − T0Sgen

This gives:

T0Sgen = m10(�11 − �10) + (m10 − m11)(�12 − �11) − WIPT


Sgen = m10(s10 − S11) + (m10 − m11)(s11 − s12)

inable

I

�

0

T

I

�

−

I

�

−

I

�

h

0

T



˙destroyed = T0Sgen = T0[m10(s10 − s11) + (m10 − m11)(s11 − s12)]


II,LPT = 1 − Idestroyed

m10(�11 − �10) + (m10 − m11)(�12 − �11)

= WLPT

m10(�11 − �10) + (m10 − m11)(�12 − �11)

(g) Condenser sub systemThe exergy balance for the condenser is give by:

= m12(�12 − �1) −n∑

k=1

(1 − T0

Tk

)Qk − T0Sgen

This gives:

0Sgen = m12(�12 − �1) −n∑

k=1

(1 − T0

Tk

)Qk


˙destroyed = T0Sgen = [{m12(h12 − h1)} − T0{m12(s12 − s1)}]

−n∑

k=1

(1 − T0

Tk

)Qk


II,Con = 1 − Idestroyed

m12(�12 − �1)

Pump sub system(h) The exergy balance for the low pressure pump is give by:

WLPP = m1(�1 − �2) − T0Sgen


˙destroyed = T0Sgen = m1(�1 − �2) + WLpp = m1T0(s2 − s1)


II,LPP = 1 − Idestroyed

WLpp= m1(�2 − �1)

WLPP

(i) The exergy balance for the High pressure pump is give by:

WHPP = m5(�5 − �6) − T0Sgen


˙destroyed = T0Sgen = m5(�5 − �6) + WHPP = m5T0(s6 − s5)


II,HPP = 1 − Idestroyed

WHPP= m5(�6 − �5)

WHPP

Feed water heater sub system(j) The exergy flow equation for the low presure feed water

eater (LPH) system becomes:

= m11(�11 − �13) − m2(�3 − �2) − T0Sgen

This gives:


Energy Reviews 15 (2011) 1857–1872 1869


Idestroyed = T0Sgen


�II,LPH = 1Idestroyed

m11(�11 − �13)= m2(�3 − �2)

m11(�11 − �13)

(k) Dearetor sub systemIt is an adiabatic mixing chamber where a hot streams 9 is mixed

with a cold stream 4, forming a mixture 5, the exergy supplied isthe sum of the exergies of the hot and cold streams, and the exergyrecovered is the exergy of the mixture. The exergy flow equationfor the dearetor system becomes:

0 = m9�9 + m4�4 − m5�5 − T0Sgen

where m5 = m9 + m4. This gives:

T0Sgen = m9�9 + m4�4 − m5�5


Sgen = m5s5 − m9s9 − m4s4


Idestroyed = T0Sgen = T0[m5s5 − m9s9 − m4s4]


Idestroyed = T0Sgen


�II,Der = 1 − T0Sgen

m9�9 + m4�4= m5�5

m9�9 + m4�4

Expansion valve(l) The exergy flow equation for the expansion valve (EXP)

becomes:

0 = m13(�13 − �14) − T0Sgen

This gives:

T0Sgen = m13[(h13 − h14) − T0(S13 − S14)]


Idestroyed = T0Sgen



m13(�13 − �14)

3.4. Results and discussion

Sue and Chuang [15] performed exergy analysis based on loadvariation. The exergy loss at 50% load is three times that of 100% loaddue to the lower steam pressure in the HRSG. Therefore, the plantselected combined cycle plant operating efficiency at 100% load is2.4% higher than at 50% load. Khaliq and Kaushik [17] presentedsecond-law efficiency of gas fire thermal power plant varying thenumber of reheat process and compression ratio in gas turbine.The first-law efficiency of the adiabatic turbine increases with theincrease in pressure ratio. The second-law efficiency decreases
with the pressure ratio, but increases with the cycle temperatureratio since a greater proportion of the available work lost at thehigher temperature may be recovered. The exergy destruction inthe reheat turbine increases with the pressure ratio, the numberof reheat stages and the pressure drop in each re-heater first-law


e cons

atisorccttwewdupttopici6iweisa

TE

Fig. 6. Exergy destructions per mole of methan

nd second-law efficiencies of the combined cycle increases upo the pressure ratio of 32, and then they start decreasing withncreases in the pressure ratio. But it is interesting to note that theecond-law efficiency of the combined cycle is greater than thatf the first-law efficiency for same pressure-ratio. If the pressureatio is too low, then the gas-turbine cycle and combined-cycle effi-iencies and their specific work-outputs drop, whereas the steamycle work-output increases due to the high gas-turbine exhaustemperature. If the pressure ratio is too high, the compressor andurbine works increase but their difference, the net gas-turbineork output drops. Franco and Russo [33] analyzed the heat recov-

ry steam generator (HRSG), as a first step in the analysis of thehole plant. They handle this problem adopting both a thermo-ynamic and a thermoeconomic objective function instead of thesual pinch point method. Thermodynamic optimization has theurpose to diminish energy losses, expressed on exergy basis, whilehe aim of the thermoeconomic optimization is the minimiza-ion of the cost function associated with the system/plant, sumf the cost of exergy inefficiencies and the cost of the HRSG. Pro-osed methods have been applied to some HRSG configurations,

ncluding some present commercial plants. The results of the appli-ation of the thermoeconomic optimization lead to a meaningfulncrease of the thermal efficiency of the plant that approaches the0%. Dincer et al. [34] reported energy and exergy assessments of

ntegrated power generation using solid oxide fuel cells (SOFCs)
ith internal reforming and a gas turbine cycle. The other main
xergy destruction is attributable to electrochemical fuel oxidationn the SOFC. The energy and exergy efficiencies of the integratedystem reach 70–80%, which compares well to the efficiencies ofpproximately 55% typical of conventional combined-cycle power

able 2nergy and exergy analyses results for four different cogeneration systems.

S. no Parameters Steam co-gener

1 Inlet temperature of hot fluid to heater (◦C) 2492 Exit temperature of hot fluid from heater (◦C) 2123 Inlet temperature of water to heater (◦C) 504 Exit temperature of water from heater (◦C) 905 Net power output, wnet (kW) 10,0006 Heating supplied, Qheat (kW) 13,5007 Exergy input to the plant � in(kW) 50.9608 Total exergy destruction in the plant, Idr (kW) 39,2009 Energy efficiency, �cogen (%) 47.810 Exergy efficiency, �cogen (%) 23.1

umed for the devices in the integrated system.

generation systems. Variations in the energy and exergy efficienciesof the integrated system with operating conditions are provided,showing, for example, that the SOFC efficiency is enhanced if thefuel cell active area is augmented. The SOFC stack efficiency can beenhanced by reducing the steam generation while increasing thestack size. Fig. 6 shows exergy destructions per mole of methaneconsumed for the devices in the integrated system. Ertesvag et al.[35] presented a concept for natural-gas (NG) fired power plantswith CO2 capture was investigated based on the exergy analysis.Natural gas was reformed in an auto-thermal reformer (ATR), andthe CO2 was separated before the hydrogen-rich fuel was used ina conventional combined-cycle (CC) process. The main purposeof the study was to investigate the integration of the reformingprocess and the combined cycle. An increase of the turbine in lettemperature (TIT) from 1250 to 1350 ◦C and1450 ◦C increased thenet electric-power production to 50.6% and 52.2%, respectively, ofthe NG for the plant with reforming and CO2 capture. The corre-sponding results for the conventional combustion chamber (CC)plant were 60.2% and 61.0% respectively [36]. For the plant withreforming and CO2 capture, a combination of 1450 ◦C TIT andATR product-feed heat exchange gave a net electric-power pro-duction of 53.3% of the NG. Kanoglu and Dincer [39] studied theperformance assessment of various cogeneration systems throughenergy and exergy efficiencies. The cogeneration plants consideredinclude steam-turbine system, gas-turbine system, diesel-engine
system, and geothermal system, and the results of the analysisare given in Table 2. Reddy and Mohamed [11] determined gasturbine main combustion chamber as the major source of exergydestruction rate. The exergy destruction rate in the main combus-tion chamber is found be very high as compared to other parts
ation Gas-turbinecogeneration

Diesel-enginecogeneration

Geothermalcogeneration

303 400 100211 111 8

50 50 5090 90 90

10,000 20,000 10,00013,500 13,500 13,50052,000 45,620 26,65040,240 23,860 14,885

46.8 78.2 16.122.6 47.7 44.1

S.C. Kaushik et al. / Renewable and Sustainable Energy Reviews 15 (2011) 1857–1872 1871

Fp(

oipnTgt

4

fn[5o[cptbpatFt6rw3ap5ibp

asdi

in coal based thermal power plant and combustion chamber in

ig. 7. Variation of (a) energy efficiency and (b) exergy efficiency with boiler tem-erature for various boiler pressures ((�) 10 MPa, (�) 11 MPa, (�) 12 MPa, (×) 13 MPa,*) 14 MPa, (�) 15 MPa) for the Ghazlan Power Plant.

f the system. Whereas the higher pressure ratio, results in anncrease in exergy destruction rate in the gas turbine cycle com-onents [40,41]. For the same pressure ratio, the combined cycleet work output increases with higher turbine inlet temperatures.he exergy destruction rate in the combustion chambers and theas turbine cycle components reduces with higher gas turbine inletemperature.

. Improvement potentials

Dincer and Rosen [46] presented exergy improvement methodsor coal fired thermal power plants. They have taken thermody-amic parameters from the Ghazlan Power Plant in Saudi Arabia42]. With the boiler pressure of 12.5 MPa and the temperature is10 ◦C. while the condenser pressure of 50.8 mmHg, temperaturef 38.4 ◦C. Whereas the regenerator pressure of 132 kPa. Habib et al.43] have indicated that for maximum efficiency of a single-reheatycle, the reheat pressure should be approximately 19% of the boilerressure. In the simulation they varied the inlet temperature ofhe high-pressure turbine (or the outlet temperature of the boiler)etween 400 ◦C and 600 ◦C with an interval of 10 ◦C. For each tem-erature, the pressure has been changed from 10 to 15 MPa withn interval of 1 MPa. The pressure at the inlet of the low-pressureurbine was taken to be 20% that of the high-pressure turbine. Inig. 7a and b show the energy and exergy efficiency has been plot-ed against the boiler temperature for values between 400 ◦C and00 ◦C, for different boiler pressures of 10, 11, 12, 13, 14 and 15 MPaespectively. All energy efficiency profiles increase almost linearlyith the boiler temperature and the energy efficiencies vary from

8% to 45%. However, the exergy efficiency varies between 52.5%nd 60%. Energy and exergy efficiency are plotted against boilerressure in Fig. 8a and b for three different temperatures (400 ◦C,00 ◦C and 590 ◦C). Although both energy and exergy efficiencies

ncrease slightly with increasing boiler pressure. The increase muste weighed against the added cost of equipment to increase theressure.

The maximum energy efficiency for the three curves occurs at
boiler pressure of 14 MPa. Therefore, this pressure can be con-
idered as a thermodynamic optimum for such a cycle under theesign conditions. Although the difference between the profiles

s larger at lower temperatures like 400 ◦C, they approach each

Fig. 8. Variation of (a) energy efficiency and (b) exergy efficiency with boiler pres-sure for various boiler temperatures ((�) 400 ◦C, (�) 500 ◦C, (�) 590 ◦C)) for theGhazlan Power Plant.

other as temperature increases and tend to match near 600 ◦C.Kelly et al. [47] analyzed the irreversibilities (exergy destruction)within a component of an energy conversion system emphasizingthat the irreversibility associated with a component can be rep-resented in two parts. The first part depends on the inefficienciesof the considered component while the second part depends onthe system structure and the inefficiencies of the other compo-nents of the overall system. Thus, the exergy destruction occurringwithin a component can be split into two parts: (a) endogenousexergy destruction exclusively due to the performance of the com-ponent being considered and (b) exogenous exergy destructioncaused also by the inefficiencies within the remaining componentsof the overall system. The paper discussed four different approachesdeveloped by the author [48] for calculating the endogenous partof exergy destruction as well as the approach based on the struc-tural theory. The advantages, disadvantages and restrictions forapplications associated with each approach have been presented.Kotas [32] explained about the mismatching of heat capacities ofheat transfer media, considering the heat transfer taking placein a parallel-flow mode and/or when the heat capacities of thestreams are mismatched in a counter-flow heat exchanger. Evenwhen the temperature difference is very small at one end of theheat exchanger, there will still be appreciable irreversibility ratedue to heat transfer over a finite temperature difference at otherpoints in the heat exchanger. This type of intrinsic irreversibility isassociated with the particular physical configuration of the plant[37,38].

5. Conclusions

Exergy analysis is shown in this article to be able to helpunderstand the performance of coal fired, gas fired combinedcycle thermal power plants and identify design possible efficiencyimprovements. It gives logical solution improving the power pro-duction opportunities in thermal power plants [44,45]. By theexergy analysis we can conclude that main energy loss in boiler

gas fired combined cycle thermal power plant. Of course, in everyplant component such as a boiler, combustion chamber there issome intrinsic irreversibility which cannot, owing to the presentstate of technological development, be eliminated. In addition,

1 inable

ewmpbdHie

R

[

[

[

[

[

[

[

[

[

[

[

[

[

[

[

[

[

[

[

[

[[

[

[

[

[

[

[

[

[

[

[

[

[

[

[

[


xergy methods are useful in assessing which improvements areorthwhile, and should be used along with other pertinent infor-ation to guide efficiency improvement efforts for steam power

lants. Of course, Efficiency of some plant components is improvedy increasing their size. For example, heat exchangers of a givenesign perform better when the heat transfer areas are increased.owever, this involves extra cost and hence there is a limit-

ng size beyond which further increase would not be justifiedconomically.

eferences

[1] Hasan HE, Ali VA, Burhanettin, Ahmet D, Suleyman HS, Bahri S, Ismail T, CengizG, Selcuk A. Comparative energetic and exergetic performance analyses forcoal-fired thermal power plants in Turkey. International Journal of ThermalSciences 2009;48:2179–86.

[2] Aljundi Islam H. Energy and exergy analysis of a steam power plant in Jordan.Applied Thermal Engineering 2009;29:324–8.

[3] Datta A, Sengupta S, Duttagupta S. Exergy analysis of a coal-based 210 MWthermal power plant. International Journal of Energy Research 2007;31:14–28.

[4] Naterer GF, Regulagadda P, Dincer I. Exergy analysis of a thermal powerplant with measured boiler and turbine losses. Applied Thermal Engineering2010;30:970–6.

[5] Rosen MA. Energy- and exergy-based comparison of coal-fired and nuclearsteam power plants. International Journal of Exergy 2001;3:180–92.

[6] Ganapathy T, Alagumurthi N, Gakkhar RP, Murugesan K. Exergy analysis ofoperating lignite fired thermal power plant. Journal of Engineering Science andTechnology Review 2009;2:123–30.

[7] Zubair SM, Habib MA. Second-law-based thermodynamic analysis ofregenerative-reheat Rankine-cycle power plants. Energy 1992;17:295–301.

[8] Reddy BV, Butcher CJ. Second law analysis of a waste heat recovery basedpower generation system. International Journal of Heat and Mass Transfer2007;50:2355–63.

[9] Suresh MVJJ, Reddy KS, Ajit KK. Energy and exergy analysis of thermal powerplants based on advanced steam parameters. In: National conference onadvances in energy research. India: IITB; 2006.

10] Oktay Z. Investigation of coal-fired power plants in Turkey and a case study:can plant. Applied Thermal Engineering 2009;29:550–7.

11] Reddy BV, Mohamed K. Exergy analysis of natural gas fired combined cyclepower generation unit. International Journal of Exergy 2007;4:180–96.

12] Srinivas T, Gupta AVSSKS, Reddy BV. Performance simulation of 210 MW natu-ral gas fired combined cycle power plant. International Journal of Energy, Heatand Mass Transfer 2007;29:61–82.

13] Can Gulen S, Smith WSR. Second law efficiency of the rankine bottoming cycleof a combined cycle power plant. International Journal of Engineering for GasTurbines and Power 2010;132:1–10.

14] Datta A, Ganguly R, Sarkar L. Energy and exergy analyses of an externally firedgas turbine (EFGT) cycle integrated with biomass gasifier for distributed powergeneration. Energy 2010;35:341–50.

15] Sue DC, Chuang CC. Engineering design and exergy analyses for combustion gasturbine based power generation system. Energy 2004;29:1183–205.

16] Bilgen E. Exergetic and engineering analyses of gas turbine based cogenerationsystems. Energy 2000;25:1215–29.

17] Khaliq A, Kaushik SC. Second-law based thermodynamic analysis ofBrayton/Rankine combined power cycle with reheat. Applied Energy2004;78:179–97.

18] Woudstra N, Woudstra T, Pirone A, van der Stelt T. Thermodynamic eval-uation of combined cycle plants. Energy Conversion and Management
2010;51:1099–110.
19] Cengel YA, Michael A. Thermodynamics an engineering approach. New Delhi:Tata McGraw Hill; 2006.

20] Kotas TJ. Exergy criteria of performance for thermal plant: second of two paperson exergy techniques in thermal plant analysis. International Journal of Heatand Fluid Flow 1980;2:147–63.

[

[

Energy Reviews 15 (2011) 1857–1872

21] Moran MJ, Shapiro HN. Fundamentals of engineering thermodynamics. NewYork: John Wiley; 1988.

22] Som SK, Mondal SS, Dash SK. Energy and exergy balance in the process ofpulverized coal combustion in a tubular combustor. Journal of Heat Transfer,Transactions of the ASME 2005;127:1322–33.

23] Poredos A, Kitanovski A. Exergy loss as a basis for the price of thermal energy.Energy Conversion & Management 2002;43:2163–73.

24] Dincer I, Rosen MA. Effect of varying dead-state properties on energy andexergy analyses of thermal systems. International Journal of Thermal Sciences2004;43:121–33.

25] Ganguly R, Ray TK, Datta A, Gupta A. Exergy-based performance analysis forproper O&M decisions in a steam power plant. Energy Conversion and Man-agement 2010;51:1333–44.

26] Saidur R, Ahamed JU, Masjuki HH. Energy exergy and economic analysis ofindustrial boilers. Energy Policy 2010;38:2188–97.

27] Dincer I, Rosen MA. Thermoeconomic analysis of power plants an applicationto a coal fired electrical generating station. Energy Conversion & Management2003;44:2743–61.

28] Norio A, Hiroshi T, Kunihiko M, Takefumi N. Exergy analysis on combustion andenergy conversion processes. Energy 2005;30:111–7.

29] Cengel YA. Heat mass transfer a practical approach. New Delhi: Tata McGrawHill; 2006.

30] Nag PK. Power plant engineering. New Delhi: Tata McGraw Hill; 2007.31] Wu C. Termodynamics and heat powered cycles: a cognitive engineering

approach. New York: Nova Science Publishers, Inc; 2007.32] Kotas TJ. The exergy method of thermal power analysis. London: Butterworths;

1985.33] Franco A, Russo A. Combined cycle plant efficiency increase based on the

optimization of the heat recovery steam generator operating parameters. Inter-national Journal of Thermal Sciences 2002;41:843–59.

34] Dincer I, Rosen MA, Zamfirescu C. Exergetic performance analysis of a gas tur-bine cycle integrated with solid oxide fuel cells. Journal of Energy ResourcesTechnology 2009;13:01–11.

35] Ertesvag IS, Kvamsdal HM, Bolland O. Exergy analysis of a gas-turbinecombined-cycle power plant with precombustion CO2 capture. Energy2005;30:5–39.

36] Kamate SC, Gangavati PB. Exergy analysis of cogeneration power plants in sugarindustries. Applied Thermal Engineering 2009;29:1187–94.

37] Horlock JH, Young JB, Manfrida G. Exergy analysis of modern fossil-fuel powerplants. Journal of Engineering for Gas Turbines and Power 2000;122:1–7.

38] Sohn JL, Song TW, Kim JH, Kim TS, Ro ST. Exergy based performance analysisof the heavy duty gas turbine in part-load operating conditions. InternationalJournal of Exergy 2002;2:105–12.

39] Kanoglu M, Dincer I. Performance assessment of cogeneration plants. EnergyConversion and Management 2009;50:76–81.

40] Srinivas T, Gupta AVSSKS, Reddy BV. Thermodynamic equilibrium model andexergy analysis of a biomass gasifier. Journal of Energy Resources Technology2009;131:1–7.

41] Reddy BV, Alaefour IE. Performance simulation of a natural gas fired combinedcycle power generation system. In: 19th national & 8th ISHMT-ASME heat andmass transfer conference. India: JNTU; 2008.

42] Al-Bagawi JJ. Energy and Exergy Analysis of Ghazlan Power Plant. M.Sc. Thesis,Mechanical Engineering Department, King Fahd University of Petroleum andMinerals, Dhahran, Saudi Arabia; 1994.

43] Habib MA, Said SAM, Al-Bagawi JJ. Thermodynamic performance analysis ofthe Ghazlan power plant. Energy 1995;20:1121–30.

44] Chen L, Li Y, Sun F, Wu C. Power optimization of open-cycle regenerator gas-turbine power-plants. Applied Energy 2004;78:199–218.

45] Reddy BV, Srinivas T, Gupta AVSSKS, Nag PK. Parametric analysis of a coalbased combined cycle power plant. International Journal of Energy Research2006;30:19–36.

46] Dincer I, Rosen MA. Exergy energy, environment and sustainable development.Elsevier; 2007.

47] Kelly S, Tsatsaronis G, Morosuk T. Advanced exergetic analysis: approaches forsplitting the exergy destruction into endogenous and exogenous parts. Energy2009;34:384–91.

48] Tsatsaronis G, Park M-H. On avoidable and unavoidable exergy destructionsand investment costs in thermal systems. Energy Conversion & Management2002;43:1259–70.

Documents

Articulo - Analisis de Energia y Exergia de Plantas de Potencia