Energy Conversion and Management 81 (2014) 282–289

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Energy Conversion and Management

journal homepage: www.elsevier .com/locate /enconman

Thermodynamic performance analysis and algorithm modelof multi-pressure heat recovery steam generators (HRSG)based on heat exchangers layout

http://dx.doi.org/10.1016/j.enconman.2014.02.0600196-8904/� 2014 Elsevier Ltd. All rights reserved.

⇑ Corresponding author. Tel.: +86 13989882228; fax: +86 571 87951058.E-mail address: wzhong@zju.edu.cn (W. Zhong).

Hongcui Feng, Wei Zhong ⇑, Yanling Wu, Shuiguang TongInstitute of Thermal Science and Power System, Zhejiang University, Hangzhou 310027, China

a r t i c l e i n f o

Article history:Received 4 November 2013Accepted 24 February 2014Available online 15 March 2014

Keywords:Heat recovery steam generator (HRSG)Heat exchangerLayout analysisThermodynamic performance

a b s t r a c t

Changes of heat exchangers layout in heat recovery steam generator (HRSG) will modify the amount ofwaste heat recovered from flue gas; this brings forward a desire for the optimization of the design ofHRSG. In this paper the model of multi-pressure HRSG is built, and an instance of a dual pressure HRSGunder three different layouts of Taihu Boiler Co., Ltd. is discussed, with specified values of inlet temper-ature, mass flow rate, composition of flue gas and water/steam parameters as temperature, pressure etc.,steam mass flow rate and heat efficiency of different heat exchangers layout of HRSG are analyzed. Thisanalysis is based on the laws of thermodynamics and incorporated into the energy balance equations forthe heat exchangers. In the conclusion, the results of the steam mass flow rate, heat efficiency obtainedfor three heat exchangers layout of HRSGs are compared. The results show that the optimization of heatexchangers layout of HRSGs has a great significance for waste heat recovery and energy conservation.

� 2014 Elsevier Ltd. All rights reserved.

1. Introduction

In recent years, with soaring energy price and increasingdemand for reducing fuel consumption, much attention has beenpaid to the utilization of industrial waste heat. Different gradesof waste heat are commonly available in metallurgy, oil, petro-chemicals and other industries. The main carrier of waste heat re-sources is sensible heat in the flue gas, and the most frequentapproach for recovery is to produce steam that can be used directlyor further used to generate electricity. The design of HRSGs is orga-nized at three levels: first it put forward the overall strategy forheat recovery, which enables to obtain pressure levels and themain operating parameters of the HRSG; the second step involvesa general layout to meet the process requirements, including thelayout of heat exchangers and the net absorbed heat of each heatexchanger; the third step leads to the detailed design of the geo-metric variables of the heat exchangers, such as tube types anddiameters, the number of tubes per row. The main goal is to in-crease heat efficiency and decrease the equipment cost with theprerequisite of guaranteeing safety and reliability. For specific val-ues of the flue gas parameters, many different choices are possibleregarding the heat recovery scheme and the general layout of heat

exchangers. Therefore, the effects of heat exchangers layout are ofgreat importance to optimize the utilization of waste heatresources.

A lot of efforts have been dedicated to the analysis and optimi-zation of HRSG thermodynamic performance. Ahmadi et al. [1]modeled the comprehensive thermodynamic modeling of a dualpressure HRSG. They carried out a multi-objective optimizationto find the best design parameters for that HRSG. Bassily [2,3]modeled a dual and a triple pressure reheat Combined Cycle PowerPlant (CCPP) for changes of the minimum pinch point temperaturedifference, temperature difference of the superheat approach, tem-perature and pressure of the steam turbine, gas outlet temperatureetc. Reddy et al. [4] used non-dimensional operating parameters toanalyze the entropy generation and to calculate the entropy gener-ation number for single pressure HRSG heat exchangers. Based onthe second law of thermodynamics, Butcher et al. [5] presented theeffects of pinch point temperature difference and flue gas compo-sition on the entropy generation rate and the second law efficiency.Valdes et al. [6–8] proposed a method for thermo-economic opti-mization of combined cycle gas turbine power plants based onthe application of influence coefficients and genetic algorithm.Sanjay [9] investigated the effect of HRSG configuration on exergydestruction of bottoming cycle components and concluded that thedistribution of exergy destruction is sensitive to the type ofbottoming cycle configuration. Woudstra et al. [10] performeda thermodynamic evaluation of CCPP with different steam

Nomenclature

g gass steami the ith heat exchangerj the jth pressure leveli inleto outletMg flue gas mass flow rate (kg/s)Ms total steam mass flow rate of public economizer (kg/s)Mj

s steam mass flow rate of the jth pressure level (kg/s)cpg specific heat at constant pressure (kJ/kg K)hs working fluid enthalpy (kJ/kg)P0 ambient pressure (MPa)Ps working fluid pressure (MPa)DTj

pp pinch point temperature difference of the jth pressurelevel (K)

DTjsatw water saturation temperature of the jth pressure level (K)

DTjgo;B gas outlet temperature of the evaporator of the jth pres-

sure level (K)DTmin minimum temperature difference (K)DTi temperature difference of the ith heat exchanger (K)DTmin,i the minimum temperature difference of the ith heat ex-

changer (K)Tg gas outlet temperature at the outlet of the economizer

(K)

Tacid acid dew point (K)T0 ambient temperature (K)gh efficiency of HRSGQ heat rate (kJ/Nm3)

Subscripts and abbreviationsHRSG heat recovery steam generatorHPS high pressure superheaterHPB high pressure evaporatorHPE high pressure economizerLPS low pressure superheaterLPB low pressure evaporatorLPE low pressure economizerECO economizerEVA evaporatorSH superheaterPH preheaterPECO public economizerRH reheaterHE heat exchangePPTM method of pinch point temperature differenceGOTM method of gas outlet temperatureDSA dichotomous search algorithm

Fig. 1a. Schematic diagram of dual pressure HRSG 1.

H. Feng et al. / Energy Conversion and Management 81 (2014) 282–289 283

bottoming cycles. The evaluation showed that the increasing num-ber of pressure levels of steam generation will reduce the lossesdue to heat transfer in the HRSG. In reference [11], a combinedpower cycle with HRSG is analyzed. Through exergy analysis, theexergy of the exhaust streams and the irreversibility of each com-ponent in the cycle are determined. In this paper [12], it is shownthat important relationships among optimal objective functionsand decision variables can be discovered consequently. Casarosaet al. [13] studied thermodynamic optimization based on the min-imization of the total HRSG cost, after the reduction to a commonmonetary base of the costs of exergy losses and of installation. In[14] a CCPP with a supplementary firing system is analyzedthrough energy and exergy, the optimal design of operating param-eters of the plant is then performed by defining an objective func-tion and applying a generic algorithm (GA) type optimizationmethod. This paper [15] shows a methodology to achieve thermo-economic optimizations of CCGT power plants taking into accountthe frequent off-design operation of the plant. In [16] a completeeconomic and thermodynamic study for dual pressure, triple pres-sure with and without reheat has been reported. Therefore, theyinvestigated the effect of the pressure levels of steam generationin HRSG on exergy efficiency of combined cycle. Mohagheghi andShayegan [17] combined with the genetic algorithm calculatedthe optimal thermodynamic performance conditions for HRSGs.Ahmadi et al. [18,19] performed the exergoenvironmental optimi-zation of a CHP system, they showed that reducing the irreversibil-ity of an HRSG increases the steam cycle efficiency due toincreasing the produced steam temperature. In this study [20], acomprehensive thermodynamic modeling of a dual pressure CCPPis performed, an optimization study to find the best design param-eters is carried out. Behbahani-nia et al. [21] considered a smallcogeneration system including a gas micro turbine and a fire tubeHRSG, the results show that the thermodynamic optimization doesnot lead to major improvement of the total cost of the HRSG due todecrease in the pinch point. In conclusion, the results show that theuse of several pressure levels in HRSGs increases the powerproduction in the steam cycle. The researches listed above mainly

focused on the analysis and optimized design of HRSG operatingparameters, and there is limited research on heat exchangerslayout.

This article presents a general model for analyzing the thermo-dynamic performance of a multi-pressure HRSG based on heatexchangers layout, in which the minimum temperature differenceof each heat exchanger will replace the constraint of pinch pointtemperature difference. Then, examples for dual pressure HRSGsare analyzed.

2. Problem model

As shown in Fig. 1a, a dual pressure HRSG model is considered,consisting of six heat exchangers (HPE, HPB, HPS, LPE, LPB, LPS). Itis assumed that the HRSG is at steady state, the heat transfer to thesurrounding is negligible, and there is no extraneous heat loss, thereference ambient conditions for air is P0 = 0.101 MPa andT0 = 293.15 K. The flue gas inlet temperature, the mass flow rate,the flue gas composition, the water/steam temperature and thepressure are given. According to the first law of thermodynamics,energy balance equations for each heat exchanger are established,and the T–Q profile is shown in Fig. 1b.

In practical cases, the heat exchangers layout will change due tosome factors. For example, a radiant heat exchanger can be placed

Fig. 1b. The T–Q profile of dual pressure HRSG 1.

Fig. 2a. Schematic diagram of dual pressure HRSG 2.

Fig. 2b. The T–Q profile of dual pressure HRSG 2.

Fig. 3b. The T–Q profile of dual pressure HRSG 3.

284 H. Feng et al. / Energy Conversion and Management 81 (2014) 282–289

in a high temperature zone, and evaporator would be placed beforesuperheater to protect it from high temperature, as shown inFig. 2a, its T–Q profile is shown in Fig. 2b. Besides, to decreasethe gas outlet temperature of multi-pressure HRSG, the feed waterflow rate of LPE could be the sum of the feed water flow rate at dif-ferent pressure levels, as shown in Fig. 3a, and the T–Q profile isshown in Fig. 3b. As depicted in Fig. 2b, when the flue gas and

Fig. 3a. Schematic diagram of dual pressure HRSG 3.

the water/steam parameters are constant, different heat exchang-ers layout brings forward changes in DTi, so that results in thechanges of DTmin,i and different thermodynamic performance.Therefore, an optimized design of HRSG heat exchangers layoutis of great importance for maximizing waste heat recovery.

3. Model equations

3.1. Control equations of the unit model

The control equation is the fundamental equation of any heatbalance calculation. Here a general calculation model for thermo-dynamic performance of a m-pressure HRSG composed of n heatexchangers is built. For a given flue gas inlet temperature, massflow rate, composition of flue gas, water/steam temperature andpressure, it follows that:

(1) In general, the heat exchanger types include: ECO, EVA, SH,and also RH and PH. Moreover, in multi-pressure HRSG,the feed water flow rate of LPE is the sum of feed water flowrate in different pressure levels and LPE is a publiceconomizer.

(2) There are two meanings of heat exchanger number: for am-pressure (1, 2, . . .j, . . .m) HRSG with n heat exchangers,on the flue gas side and on its flow direction, heat exchang-ers could be numbered as: 1, 2, . . .i, . . .n; on water/steamside and on its flow direction, assume there are s heatexchangers under jth pressure level, then they can be num-bered as: 1, 2, . . .k, . . .s. Taking Fig. 1a as example, it is a dualpressure (m = 2) HRSG with 6 heat exchangers (n = 6), on theflue gas side, it can be numbered as i = 4 and on water/steamside as j = 2, k = 1.

(3) As shown in Fig. 4, heat balance calculation model could bebuilt for each heat exchanger. Mg is given, cpg is assumed asconstant in given temperature zone, based on the first law ofthermodynamics [22], by applying the energy balance forflue gas and water/steam in each heat exchanger of theHRSG, the flue gas temperature and water/steam propertieswill be calculated by the following equations:

Fig. 4. Model of the ith heat exchanger.

Fig. 5. Model of a public low pressure economizer.

H. Feng et al. / Energy Conversion and Management 81 (2014) 282–289 285

MgcpgðTgi;i � Tgo;iÞ ¼ Mjsðh

jso;k � hj

si;kÞ ð1Þ

n equations could be built for n heat exchangers.Especially, in the case with public economizer, as described in

Fig. 5, the energy balance equation on flue gas side is given by:

MgcpgðTgi;pub � Tgo;pubÞ ¼ M1s ðh

1so;1 � hsi;pubÞ þ � � � þMj

sðhjso;1 � hsi;pubÞ

þ � � � þMms ðh

mso;1 � hsi;pubÞ ð2Þ

Ms ¼ M1s þ � � � þMj

s þ � � � þMms ð3Þ

3.2. Correlate equations

Correlate equations describe the logical relationship of flue gasand water/steam flow through HRSG heat exchangers.

On the flue gas side, the inlet temperature of ith heat exchangeris the outlet temperature of (i�1)th heat exchanger, and the outlettemperature of ith heat exchanger is the inlet temperature of(i + 1)th heat exchanger:

Tgi;i ¼ Tgo;ði�1Þ; Tgo;i ¼ Tgi;ðiþ1Þ ð4Þ

On water/steam side, under jth pressure level, the inlet temper-ature of kth heat exchanger is the outlet temperature of (k�1)thheat exchanger, and the outlet temperature of kth heat exchangeris the inlet temperature of (k + 1)th heat exchanger:

Tjso;k ¼ Tj

si;kþ1; Pjso;k ¼ Pj

si;kþ1

Tjsi;k ¼ Tj

so;k�1; Pjsi;k ¼ Pj

so;k�1

(; ð1 6 j 6 m; 1 6 k 6 sÞ ð5Þ

hjso;k¼ f ðTj

so;k;Pjso;kÞ¼hj

si;kþ1¼ f ðTjsi;kþ1;P

jsi;kþ1Þ

hjsi;k¼ f ðTj

si;k;Pjsi;kÞ¼hj

so;k�1¼ f ðTjso;k�1;P

jso;k�1Þ

8<: ;ð16 j6m; 16k6 sÞ

ð6Þ

3.3. Complementary equations

Heat balance calculation results can be obtained based on theequations above, but constraints such as acid dew point tempera-ture and pinch point temperature are not considered, so that com-plementary equation are increased to verify and validate the results.

3.3.1. The minimum temperature differenceIn HRSG, temperature difference of each heat exchanger be-

tween the flue gas inlet temperature and water/steam outlet tem-perature are different between different heat exchanger type andactual production requirements. The minimum value of each heatexchanger temperature difference between the flue gas inlettemperature and water/steam outlet temperature, which calledthe minimum temperature difference, could be set to achieve themaximum steam mass flow rate.

For evaporator under each pressure level, the minimum tem-perature difference is the difference between the flue gas outlettemperature and the feed water temperature under given pressure,called pinch point temperature difference:

DTjpp ¼ Tj

go;B � Tjsatw ð7Þ

As mentioned above, on flue gas side, the temperature differ-ence between the flue gas inlet temperature and water/steam out-let temperature would be changed with the change of heatexchanger layout. Therefore, the minimum temperature differ-ences of each heat exchanger are the constraints:

DTi ¼ ðTgi;i � Tjso;kÞP DTmin;i ð8Þ

3.3.2. Gas outlet temperatureGas outlet temperature should be higher than acid dew point

temperature to prevent corrosion of heat exchangers; gas outlettemperature is the flue gas outlet temperature of the nth heatexchanger:

Tg ¼ Tgo;n P Tacid ð9Þ

Under high flue gas temperature, gas outlet temperature couldreach a low level, when Tg is lower than Tacid, Tacid would constraintTg, in order to make Tg higher than or equal to Tacid, we assume Tg isequal to Tacid.

4. Mathematical solution

There are m steam mass flow rates Mjs and n gas outlet temper-

atures Tgo,i by the equations listed above, and the steps for calculat-ing the (m + n) unknowns are as follow:

(1) Calculation of the initial values (PPTM):

It is assumed that m pinch point temperature differences DTjpp

are the minimum temperature difference, we can get the gas outlettemperature based on Eq. (7):

Tjgo;B ¼ DTj

pp þ Tjsatw ð10Þ

For m steam outlet temperatures of evaporator and n energybalance equations, m Mj

s and (n�m) Tgo,i are calculated; and the fluegas inlet temperature Tgi,i are calculated by Eq. (4).

(2) The verification of constraints:(a) Temperature difference: DTi would be checked based Tgi,i and

Eq. (8); if it does not satisfy the constraints, the initial valuesof Mj

s would be adjusted, which is shown in step (3a) indetail; otherwise Tg would be checked.

(b) The acid dew point temperature: the gas outlet temperaturewould be checked with Eq. (9). If Tg is lower than Tacid, Tg

need to be adjusted in step (3b); otherwise the calculationprocess would complete.

(3) Adjustment of the initial value:(a) The minimum temperature difference (method of iteration):

according to the energy balance equations, the ith heatexchanger heat is equal to the kth heat exchanger underjth pressure level:

Qi ¼ Q jk ¼ MgcpgðTgi;i � Tgo;iÞ ¼ Mj

scjp;kðT

jso;k � Tj

si;kÞ ð11Þ

Assume cjp;k is specific heat at constant pressure in each heat ex-

changer; slopes of each section in T–Q profile are the reciprocal ofMj

s times cjp;k, which can be written as:

Tjso;k � Tj

si;k

� �=Q j

k ¼ 1=Mjsc

jp;k ð12Þ

In T–Q profile, when DTi is lower than DTmin,i, DTi and Tgi,i needto be increased with Eq. (8). Take Fig. 1b for example, parametersare unchanged, when the temperature difference of heat exchanger4 increases, M1

s could be decreased, so that Q3 decreases, the slope

286 H. Feng et al. / Energy Conversion and Management 81 (2014) 282–289

of Section 3 becomes bigger and it gets shorter, the slopes of LPheat exchangers all become bigger and the gas outlet temperatureincreases. We can use dichotomous search algorithm to adjust theresults, the steps are given as:

i. Assuming the flue gas to be an ideal gas and the values of cpg

the same in different temperature zone, heat balance calcu-lation would be a system of linear equations. Mj

s calculatedin step (1) would be decreased for certain percent (10%e.g.) from LP in sequence, and the new steam mass flow ratewould be:

Mjs2 ¼ ð1—10%ÞMj

s ð13Þ

ii. Plug the new steam mass flow rate into energy balanceequations to get new flue gas inlet temperature, if the resultsobtained satisfy the constraints, DSA would be used for Mj

sn:

MjsD1 ¼ ðM

js þMj

s2Þ=2 ð14Þ

Fig. 6. Flow chart of HRSG heat balance calculation.

Table 1Parameters of a dual pressure HRSG.

Flue gas inlet temperature 350Gas mass flow rate (Nm3/h) 1,000,000Heat loss (%) 0.5Dust content (g/Nm) 9Flue gas composition (%)CO2 0H2O 0N2 75.42O2 24.58

iii. Put into energy balance equations for verification, to get newflue gas inlet temperature, then verify the constraint, if theresults obtained satisfy the constraints, DSA would be usedfor Mj

sn:DSA would be used if the results calculated meet the require-

ment, or decreasing Mjsn for certain percent (10% e.g.) until it meets

the demand of constraints. If it could not meet the demand to ad-just Mj

sn, then Mjþ1s would be adjusted and the steps are the same as

adjusting Mjsn.

(b) The gas outlet temperature (GOTM):

In step (2), Tacid would be the constraint when Tg calculated islower than Tacid, we first assume Tg is equal to Tacid, then:

Mgcpg Tg1i � Tg� �

¼ M1s h1

so � h1si

� �þ � � �Mj

s hjso � hj

si

� �þ � � �Mm

s hmso � hm

si

� �ð15Þ

To ensure the amount of HP Ms, assume M1s is unknown, other

steam mass flow rates are calculated by the minimum temperaturedifference constraint. For given water/steam parameters and Tg, M1

s

could be calculated, and temperature difference under 1st pressurelevel would be changed, and the verification steps are the same asabove mentioned.

The specific steps of HRSG heat balance calculation are shownin Fig. 6.

The efficiency of HRSG is:

gHRSG ¼ ðTg1i � TgÞ=ðTg1i � T0Þ ð16Þ

5. Results and discussion

5.1. The effects of heat exchangers layout on HRSG thermodynamicperformance

Given three examples of dual pressure HRSG based on differentheat exchangers layout. As mentioned above, flue gas inlet temper-ature, mass flow rate, composition of flue gas, the water/steamtemperature and pressure in three examples are the same, otherparameters of dual pressure HRSG are shown in Table 1. Schematicdiagrams of three dual pressure HRSGs are shown in Figs. 1a–3aand corresponding T–Q profiles are Figs. 1b–3b. According to thegeneral model for thermodynamic performance of multi-pressureHRSG based on heat exchangers layout mentioned above, basedon the laws of thermodynamics, thermodynamic performance fordual pressure HRSG based on three different heat exchangerslayout are shown in Tables 2–4.

5.1.1. The effects of heat exchangers layout on HRSG steam mass flowrate

Fig. 7 shows the comparative analysis of steam mass flow ratein different pressure levels in three heat exchangers layout. Inwhich the total steam mass flow rate Ms, high pressure steammass flow rate M1

s and low pressure steam mass flow rate M2s

are all different with different heat exchangers layout: dual pres-sure HRSG with public economizer in layout 3 has the maximumM1

s , while its M2s and Ms are the smallest, steam mass flow rates of

layout 1 and layout 2 in different pressure levels are nearly thesame.

5.1.2. The effects of heat exchangers layout on HRSG efficiencyFig. 8 presents the efficiencies of HRSG in three different heat

exchangers layout HRSGs. Other parameters are given, in Tables2–4, the gas outlet temperature Tg under layout 1 is the lowest,and by Eq. (16), it has the highest efficiency; the efficiency of lay-out 2 and layout 3 are nearly the same, there are little differencesbetween them. The reason is because of the principle of cascadeutilization of energy, the layout of heat exchangers in layout 1adherence to this principle better, while in layout 2 and layout 3other factors such as the practical factors were considered; andthe small differences of heat efficiency in three different layoutsillustrate the parameters would be the main factors for heat effi-ciency of HRSG.

Table 2Gas and water/steam properties of HRSG 1.

HE-No. Tjsi;i (�C) Tj

so;i (�C) Pjs;i (MPa) Mj

s;i (kg/s) Tgi,i (�C) Tgo,i (�C) Mg (kg/s) Q (kJ/Nm3)

1 218.6 320 2.16 19.89 350 335.5 176.03 64.402 218.6 218.6 2.16 20.47 335.5 272.2 172.11 35.963 161.8 230 0.55 12.06 272.2 266.9 154.36 83.634 110 213.6 2.37 20.89 266.9 204.1 153.03 6.775 161.8 161.8 0.55 12.49 204.1 176.8 135.36 85.306 110 156.8 0.6 12.67 176.8 127.6 127.75 19.24

Table 3Gas and water/steam properties of HRSG 2.

HE-No. Tjsi;i (�C) Tj

so;i (�C) Pjs;i (MPa) Mj

s;i (kg/s) Tgi,i (�C) Tgo,i (�C) Mg (kg/s) Q (kJ/Nm3)

1 218.6 218.6 2.16 20.47 350 340.5 176.03 62.722 218.6 320 2.16 19.89 340.5 326.1 173.53 37.633 218.6 218.6 2.16 20.64 326.1 272.5 169.64 83.634 161.8 230 0.55 12.06 272.5 267.3 154.61 6.775 110 213.6 2.37 20.89 267.3 205 153.28 72.766 161.8 161.8 0.55 12.5 205 176.7 135.75 19.24

Table 4Gas and water/steam properties of HRSG 3.

HE-No. Tjsi;i (�C) Tj

so;i (�C) Pjs;i (MPa) Mj

s;i (kg/s) Tgi,i (�C) Tgo,i (�C) Mg (kg/s) Q (kJ/Nm3)

1 218.6 320 2.16 21.83 350 334.2 176.03 62.722 218.6 218.6 2.16 22.92 334.2 271.4 171.75 37.013 161.8 230 0.55 10 271.4 267 154.14 83.634 156.8 213.6 2.37 22.92 267 204.7 153.06 5.655 161.8 161.8 0.55 11.19 204.7 176.8 135.53 85.306 110 156.8 0.6 31.33 176.8 129.3 127.75 21.12

Fig. 7. Steam mass flow rate in each pressure level of three heat exchangers layout.

Fig. 8. Heat efficiency of three different heat exchangers layout.

H. Feng et al. / Energy Conversion and Management 81 (2014) 282–289 287

5.1.3. The effects of heat exchangers layout on HRSG absorbed heatrate of each heat exchanger

Figs. 9–11 depicts absorbed heat rate of each heat exchanger inthree different heat exchangers layout HRSGs. To match the mini-mum temperature difference of HRSGs heat exchangers under gi-ven water/steam parameters, absorbed heat rate would bedifferent in three different heat exchangers layout.

Therefore, in practical application, we could choose the opti-mum heat exchangers layout for HRSG to meet the demand ofpractical production.

The flue gas temperature in the above example is relatively low,according to the second law of thermodynamics, to decrease theirreversibility of the heat transfer process, the effect of heatexchangers layout is obvious; if the gas temperature is high en-ough, to get the same steam, the exergy would not be the limita-tion, so that heat exchangers layout may not have a significanteffect on HRSG thermodynamic performance. Therefore, the practi-cal conditions should be considered to decide whether the optimi-zation of heat exchangers layout is feasible.

5.2. The effects of parameters on HRSG thermodynamic performance

Except for the effect of heat exchangers layout, parameter isanother key factor on HRSG thermodynamic performance. Under

Fig. 9. Absorbed heat rate of HRSG 1.

Fig. 10. Absorbed heat rate of HRSG 2.

Fig. 11. Absorbed heat rate of HRSG 3.

Fig. 12. The T–Q profile of the increasing of HP superheated steam pressure.

Fig. 13. The T–Q profile of the increasing of HP superheated steam temperature.

288 H. Feng et al. / Energy Conversion and Management 81 (2014) 282–289

given HRSG layout, with the variation of flue gas temperature, gasmass flow rate, flue gas composition, superheated steam pressureand temperature, feed water pressure and temperature in HRSG,the quantity of waste heat recovery in flue gas is different. TakeHP superheated steam pressure and temperature for example, toanalyze their effects on HRSG thermodynamic performance.

5.2.1. The effect of HP superheated steam pressure P2s

As shown in Fig. 12, other parameters are constant, when P2s in-

creases, h2s rises and M2

s decreases by Eqs. (1) and (7), and theslopes of sections ef and jk become bigger, which are shown as e0f0

and j0k; by equations of each heat exchanger and Eq. (8), M1s in-

creases and the slopes of sections ab and gh become smaller, whichare shown as a0b0 and g0h0.

5.2.2. The effect of HP superheated steam temperature T2s

As presented in Fig. 13, other parameters remain unchanged,when T2

s increases, h2s rises and M2

s decreases by Eqs. (1) and (7),and the slopes of sections ef and jk become bigger, which areshown as e0f0 and j0k; Q2 decreases for decreasing M2

s , and in T–Qprofile it is shown as ij gets shorten to i0j0; by equations of each heatexchanger and Eq. (8), M1

s increases and the slopes of sections aband gh become smaller, which are shown as a0b0 and g0h0, Q5 in-creases for increasing M1

s , and in T–Q profile it is shown as cd be-comes longer to c0d0.

6. Conclusions

In this article, a general model for analyzing the thermodynamicperformance of multi-pressure HRSG based on heat exchangerslayout is built based on the laws of thermodynamics, in whichthe minimum temperature difference of each heat exchanger isintroduced to replace the constraint of pinch point temperaturedifference, and three different heat exchangers layout of dual pres-sure HRSGs models are analyzed, the heat balance calculation re-sults prove that the selection of heat exchangers layout hassignificant effects on the thermodynamic performance of HRSGs,so that choosing the optimal layout of HRSGs based on the practice

H. Feng et al. / Energy Conversion and Management 81 (2014) 282–289 289

requirements will of great significance for waste heat recovery forbetter energy conservation. In the following work, parameters andlayout could be considered together, and the genetic algorithmcould be used to find the optimum solution of the whole systemto save more energy.

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