iii 21 performance analysis of different working fluids for use in orc

Performance analysis of different working fluidsfor use in organic Rankine cyclesP J Mago, L M Chamra, and C SomayajiDepartment of Mechanical Engineering, Mississippi State University, Mississippi, USA

The manuscript was received on 26 September 2006 and was accepted after revision for publication on 9 November 2006.

DOI: 10.1243/09576509JPE372

Abstract: This article presents a second-law analysis for the use of organic Rankine cycle (ORC)to convert waste energy to power from low-grade heat sources. The organic working fluids wereselected to investigate the effect of the fluid boiling point temperature on the performance ofORCs. The working fluids under investigation are R134a, R113, R245ca, R245fa, R123, isobutane,and propane, with boiling points between 243 and 48 8C. The results are compared with thoseof water under similar conditions. A combined first- and second-law analysis is performed byvarying some system operating parameters at various reference temperatures. Some of theresults demonstrate that ORC using R113 shows the maximum efficiency among the evaluatedorganic fluids for temperatures.430 K; R123, R245ca, and R245fa show the best efficiencies fortemperatures between 380 and 430 K; and for temperatures ,380 K, isobutane shows the bestefficiency. Also, it is shown that the organic-fluid boiling point has a strong influence on thesystem thermal efficiency.

Keywords: organic Rankine cycle, organic fluids, refrigerants, waste heat

1 INTRODUCTION

Several industrial processes have low-temperaturewaste heat sources that cannot be efficiently recov-ered. Owing to lack of efficient recovery methods,low-grade waste heat has generally been discardedby industry and has become an environmental con-cern because of thermal pollution [1]. A solution tothis problem is the use of organic Rankine cycles(ORCs), which can make use of low-temperaturewaste heat to generate electricity. The ORC is a ther-modynamic cycle that uses an organic working fluidto generate electricity. The working fluid is heated toboiling, and the expanding vapour is used to drive aturbine. This turbine can be used to drive a generatorto convert the work into electricity. The working-fluid vapour is condensed back into a liquid andfed back through the system. A schematic of asimple ORC configuration is shown in Fig. 1.One advantage of using ORCs instead of steam

Rankine cycles in the energy industry is that the

thermal efficiency becomes economically feasiblewhen organic fluids are used to recover waste heat,370 8C [1]. ORCs can be applied for low-tempera-ture waste heat recovery (industry), efficiencyimprovement in power stations [2], and recovery ofgeothermal and solar heat. Therefore, the use ofORC to recover waste heat is favourable in manyaspects: better and economical use of the energy aswell as reduce the emission of CO2. Some examplesof low-grade waste heat (80200 8C) are industrialwaste streams, solar heat trapped in the collectors,cooling water streams of stationary diesel engines,and the exhaust of diesel engines and biomass,among others.Some of the researchers who have investigated the

application and performance of ORCs are Hung et al.[1], Hung [3], Gurgenci [4], Yamamoto et al. [5], Leeet al. [6], Larjola [7], and Larjola et al. [8], amongothers.Another important aspect of the operation of ORCs

is the selection of the working fluid. The organicworking fluid must be carefully selected on thebasis of safety and technical feasibility. There is awide selection of organic fluids that could be usedin ORC. Performances and characteristics of different

Corresponding author: Department of Mechanical Engineering,

Mississippi State University, 210 Carpenter Engineering Building,

PO Box ME, MS 39762-5925, USA. email: [email protected]

255

JPE372 # IMechE 2007 Proc. IMechE Vol. 221 Part A: J. Power and Energy

working fluids for waste heat recovery system can befound in references [9] to [11]. Generally, a goodwork-ing fluid should exhibit low toxicity, good materialcompatibility and fluid stability limits [12], and lowflammability, corrosion, and fouling characteristics.Refrigerants are good candidates for ORC appli-cations because of their low-toxicity characteristics[13]. Another characteristic that must be consideredduring the selection of organic fluid is its saturationvapour curve. This characteristic affects the fluidapplicability, cycle efficiency, and arrangement ofassociated equipment in a power generation system[3]. The slope of the saturation curve in the Ts dia-gram depends on the type of fluid employed. A dryfluid has a positive slope; a wet fluid has a negativeslope; and an isentropic fluid has infinite largeslopes. Generally, dry and isentropic fluids arebetter working fluids for an ORC because they donot condensate after the fluid goes through the tur-bine. The comparison of the temperatureentropydiagram for dry, wet, and isentropic fluids is pre-sented in Fig. 2. The working fluids employed inthis investigation are classified as follow: R113,R123, R245ca, R245fa, and isobutane are dry fluids,whereas water, R134a, and propane are wet fluids.Isentropic fluids such as R12 and R22 were notincluded as they are being phased-out and replacedwith alternative refrigerants. Hung et al. [1] pre-sented an analysis of ORC using R134a and R113,but they did not include a second-law analysis.Hung [3] performed a second-law analysis of anORC using R113 and R123. Fluids such as propaneand isobutane have been analysed for a combinedpower and cooling cycle by Vijayaraghavan andGoswami [10]. Their research includes combinedfirst- and second-law analysis for their specific

Fig. 1 Simple configuration of ORC to produce electrical power

Fig. 2 Comparison of the working fluids: (a) isentropic,

(b) wet, and (c) dry

256 P J Mago, L M Chamra, and C Somayaji

Proc. IMechE Vol. 221 Part A: J. Power and Energy JPE372 # IMechE 2007

cycle. Although some of the fluids selected for thisinvestigation have been already analysed for differ-ent cycles, R113 and R123 are the only ones thathave been specifically evaluated for ORC from theexergy point of view. Therefore, this article presentscombined first- and second-law analysis foradditional fluids such as R245ca, R245fa, isobutane,propane, and R134a for use in ORCs. Different pro-cedures to perform the second-law analysis of powercycles can be found in references [14] and [15],among others.The objective of this article is to study the change

in thermal efficiency and irreversibility of an ORCusing different working fluids varying system-operat-ing parameters at various reference temperatures.Also, the influence of the fluid boiling point tempera-ture on the performance of ORCs is investigated.

2 ANALYSIS

The equations used to determine the cycle efficien-cies as well as the cycle irreversibility of ORC are pre-sented in this section. Using the first and second lawsof thermodynamics, the performance of an ORC canbe evaluated under diverse working conditions fordifferent working fluids. The analysis presented inthis article assumes the following: steady-stateconditions, no pressure drop in the evaporator, con-denser, and pipes, and isentropic efficiencies for theturbine and pump. A simple configuration of ORC forconverting waste heat into useful electrical power isshown in Fig. 1. As can be observed from Fig. 1,there are four different processes: process 1-2(pumping process), process 2-3 (constant-pressuretransfer of heat), process 3-4 (expansion process),and process 4-1 (constant-pressure heat transfer).

2.1 Process 1-2 (pump)

The circulation pump is the driving mechanism ofthe ORC. The pump power can be expressed as

_W p _W p,idealhp

_m(h1 h2s)hp

(1)

where _W p,ideal is the ideal power of the pump, _m theworking fluid mass flowrate, hp is the isentropic effi-ciency of the pump, and h1 and h2s the enthalpies ofthe working fluid at the inlet and outlet of the pump,respectively, for the ideal case. The actual specificenthalpy of the working fluid at the pump outlet is

h2 h1 _W p_m

(2)

The irreversibility rate for uniform flow conditionscan be expressed as

_I To dSdt To _m

Xsexit

Xsinlet

h

dssystemdt

Xj

q j

T j

#(3)

where the subscript j represents the heat transfer fordifferent reservoirs and the term (dssystem/dt) 0 forsteady-state conditions.For the pump, the irreversibility rate can be

expressed as

_Ip To _m(s2 s1) (4)

where s1 and s2 are the specific entropies of the work-ing fluid at the inlet and exit of the pump for theactual conditions, respectively.

2.2 Process 2-3 (evaporator)

This is a constant-pressure transfer of heat. Theevaporator heats the working fluid at the pumpoutlet to the turbine inlet condition, which can besaturated or superheated vapour. The heat transferrate from the evaporator into the working fluid isgiven by

_Qe _m(h3 h2) (5)

where h3 and h2 are the enthalpies of the workingfluid at the exit and inlet of the evaporator,respectively.Using equation (3), the evaporator irreversibility

rate can be determined as

_Ie To _m (s3 s2) h3 h2TH

(6)

where s3 and s2 are the specific entropies of the work-ing fluid at the inlet and exit of the evaporator,respectively, and TH is the temperature of the high-temperature heat source. This temperature is con-sidered to be equal to TH T3 DTH.

2.3 Process 3-4 (turbine)

The superheated or saturated vapour working fluidpasses through the turbine to generate the mecha-nical power. After the vapour expands, it is depressu-rized by the turbine blades. The turbine power isgiven by

_W t _W t,idealht _m(h3 h4s)ht (7)

Performance analysis of different working fluids 257


where _W t,ideal is the ideal power of the turbine, ht theturbine isentropic efficiency, and h3 and h4s theenthalpies of the working fluid at the inlet andoutlet of the turbine for the ideal case. The actualspecific enthalpy of the working fluid at the turbineexit is

h4 h3 _W t_m

(8)

The turbine irreversibility rate can be expressed as

_I t To _m(s4 s3) (9)

where s3 and s4 are the specific entropies of the work-ing fluid at the inlet and exit of the turbine for theactual conditions, respectively.

2.4 Process 4-1 (condenser)

The working fluid leaving the turbine goes through aconstant-pressure phase change process in the con-denser into a saturated liquid, rejecting latent heatinto the environment or the condenser coolant. Thecondenser heat rate, _Qc, which is the rate of latentheat rejection from the condensing working fluid,can be expressed as

_Qc _m(h1 h4) (10)

The condenser irreversibility rate can be determinedfrom equation (3) as follows

_Ic To _m (s1 s4) h1 h4TL

(11)

where s1 and s4 are the specific entropies of theworking fluid at the inlet and exit of the condenser,respectively, and TL is the temperature of thelow-temperature reservoir. This temperature isconsidered to be equal to TL T12 DTL.

2.5 Cycle efficiency

The thermal efficiency is defined as the ratio betweenthe net power of the cycle to the evaporator heatrate. It can be expressed as

hcycle _W t _W p

_Qe(12)

2.6 Total cycle irreversibly

The total irreversibility can be obtained by addingequations (4), (6), (9), and (11) as follows

_Icycle Xj

_I j _Ip _Ie _I t _Ic (13)

_Icycle _mTo h3 h2TH

h1 h4

TL

(14)

2.7 Second-law efficiency

The second-law cycle efficiency can be calculatedusing the following equation

hII _W net

_Qe(1 (TL=TH))(15)

3 RESULTS AND DISCUSSION

For the purpose of this study, seven organic fluidswithdifferent boiling points ranging from 243 to 48 8Cwere employed. These organic fluids are R134a,R113, R245ca, R245fa, R123, isobutane, and propane.The results for the different organic fluids were com-pared with those of water under similar operatingconditions. Some of the properties of the fluids usedin this investigation are presented in Table 1.Figure 3 shows the variation of the system thermal

efficiency with the turbine inlet temperature. Basi-cally, this figure shows the effect of superheating ofthe working fluid over the thermal efficiency of thecycle. The range of temperature used to analyseeach fluid varies from the saturation temperature tothe critical temperature. To generate this figure, theevaporator pressure and condenser temperaturewere kept constant at 1.5 MPa and 298 K, respect-ively. The isentropic efficiencies of the turbine andpumpwere 80 and 85 per cent, respectively, althoughthe temperature differential was kept constant at15 K for all cases. This figure illustrates that the effi-ciency of the cycle for the evaluated organic fluidsis a weak function of the turbine inlet temperature,because it remains approximately constant orslightly decreases with the increment of the turbineinlet temperature. This reflects the fact that organicfluids do not need to be superheated to increasethe cycle thermal efficiency as opposed to water,which on increasing the inlet turbine temperatureincreases the thermal efficiency. Even though onlyone evaporator pressure is shown in Fig. 3, theauthors verified that the performance of other press-ures is similar to the one presented here. Figure 3 canalso be used to analyse the influence of the fluid



boiling point temperature on the system thermalefficiency. The boiling point temperatures areshown next to each fluid in Fig. 3. It can be observedthat the fluid that shows the best thermal efficiency isR113, which has the highest boiling point among theselected fluids (47.59 8C), whereas the fluid withthe worst thermal efficiency is propane, whichhas the lowest boiling point (242.09 8C). Similartrend is observed for the remaining working fluidsselected in this investigation. Therefore, it can beconcluded that the higher the boiling point of thefluid, the better the cycle thermal efficiency. FromFig. 2, it can also be observed that R113 shows themaximum efficiency among the organic fluids fortemperatures .430 K, whereas R123, R245ca, andR245fa show the best efficiencies for temperaturesbetween 380 and 430 K. For temperatures ,380 K,isobutane shows the best efficiency, whereas wateris the best fluid when temperatures go 470 K for theanalysed conditions. This figure also illustrates thatthe selected organic dry fluids (R113, R123, R245ca,R245fa, and isobutane) show better performancethan the selected organic wet fluids (R134a andpropane). One of the reasons why dry fluids showbetter thermal efficiencies than wet fluids is becausethey do not condense after the fluid goes throughthe turbine, as opposed to wet fluids, which condenseafter the turbine. However, it is important to pointout here that independently of the fluid classification,organic fluids can be used to produce power fromlow-temperature waste heat. However, organicfluids are restricted to a small range of applications,depending on their thermodynamic conditions.Figure 4 shows the variation of the system-specific

irreversibility with the turbine inlet temperature

Table

1Propertiesoftheorganic

fluidsusedin

thisinvestigation

Fluid

Type

Molecular

weight

(kg/m

ol)

Boiling

point

(8C)

Criticalproperties

Rangeofapplication

P (MPa)

T (8C)

Density

(kg/m

3)

Minim

um

temperature

(8C)

Maximum

temperature

(8C)

Maximum

pressure

(MPa)

R134a[16]

Wet

102.03

226.07

4.059

101.06

511.9

2103.3

180.0

70

R113[17]

Dry

187.38

47.59

3.392

214.06

560.0

236.2

251.9

200

R123[18]

Dry

152.93

27.82

3.662

183.68

550.0

2107.2

326.9

40

R245ca[19,20]

Dry

134.05

25.13

3.925

174.42

523.6

273.2

226.9

60

Propane[21]

Wet

44.10

242.09

4.248

96.70

220.5

2187.3

326.9

733

Isobutane[21]

Dry

58.12

211.61

3.64

134.7

224.4

2159.6

326.9

35

R245fa

[22]

Dry

134.05

14.90

4.25

154.05

517.0

273.2

226.9

60

Fig. 3 Variation of the system thermal efficiency with

the turbine inlet temperature (Pe 1.5 MPaand Tc 298 K)



under the same conditions used to generate theresults shown in Fig. 3. It can be observed that thetotal system irreversibility increases with the incre-ment of the turbine inlet temperature for all thefluids. The results presented in this figure show theimportance to perform a second-law analysis.According to the results presented in Fig. 3, the ther-mal efficiency is approximately constant with theincrement of the turbine inlet temperature. How-ever, a combined first- and second-law analysisshows that the best case scenario is obtained whenthe fluid is operated at saturated conditions beforethe turbine. This yields the same thermal efficiencywith lower irreversibility that operating under super-heated conditions does. Figure 4 also illustrates howthe system with higher (R113) and lower (propane)thermal efficiencies present the lower and higherirreversibilities, respectively. Among all the evaluatedfluids, water shows the highest specific irreversibility.The effect of the boiling point temperature in thisfigure does not present a consistent trend, as theone shown in Fig. 3.Figure 5 illustrates the effect of the turbine inlet

temperature on the system second-law efficiency. Itcan be observed for all the fluids that the second-law efficiency decreases with the turbine inlet temp-erature. These results agree well with the results pre-sented in Fig. 4, because an increment in the systemirreversibility yields a decrease in the system second-law efficiency. For temperatures between 430 and525 K, R113 shows the best second-law efficiency;for a range of temperatures 400430 K, R123 showsthe best efficiency; R245ca and R245fa present thebest second-law efficiency for temperatures between380 and 400 K. Isobutane shows the best efficiencies

for a temperature range of 360380 K and R134a for atemperature range of 330360 K, whereas propaneshows the low second-law efficiency among all theevaluated fluids. The effect of the turbine isentropicefficiency on the system second-law efficiency wasalso evaluated. It was found that an increase in theturbine isentropic efficiency represents an increasein the second-law efficiency for all the evaluatedfluids.Figure 6 illustrates the variation of the system ther-

mal efficiency with the turbine inlet pressure whilekeeping the turbine inlet temperature at saturatedconditions. For this case, the condenser temperaturewas kept constant at 298 K, and the maximumpressure used for each fluid was the critical pressure.

Fig. 4 Variation of the system irreversibility with the

turbine inlet temperature (Pe 1.5 MPa andTc 298 K)

Fig. 5 Variation of the system second-law efficiency

with the turbine inlet temperature

(Pe 1.5 MPa and Tc 298 K)

Fig. 6 Variation of the system thermal efficiency with

the turbine inlet pressure (Tc 298 K)



Similar to Fig. 3, the isentropic efficiencies of the tur-bine and pump were 80 and 85 per cent, respectively,and the temperature differential was kept constant at15 K. The results are consistent for all the fluids,because the system thermal efficiency increaseswith the increment of the turbine inlet pressure forall of them. This can be explained since with anincrease in the inlet turbine pressure, both the network and the evaporator heat increase. However,the percentage of increase of the net work is higherthan that of increase of the evaporator heat. There-fore, the ratio of the net work to the evaporatorheat increases with the turbine inlet pressure. Similarto the results presented in Fig. 3, this figure showsthat R113 has the best performance among theorganic fluids for pressures ,3.4 MPa; R123 andR245ca present the best efficiencies for a range ofpressures between 3.4 and 3.6 MPa, and R134a pre-sents the best efficiency for a range of pressuresbetween 3.6 and 4.2 MPa. Water shows the best ther-mal efficiency under the conditions analysed in thiscase. The trend observed with the boiling pointdescribed in Fig. 2 is also consistent with the resultspresented in this figure.Figure 7 showed the total specific irreversibility

versus the turbine inlet pressure for the same con-ditions used to generate Fig. 4, for R113, R123,R245fa, and isobutane. Some of the fluids wereomitted from this figure to improve its readability.It can be observed how the irreversibility increaseswith the increment of the turbine pressure for allthe fluids. Water shows the highest irreversibilityvalues and they are not shown in the figure in orderto be able to observe the values for the remainingfluids. Among the organic fluids, isobutane and

R134a show the highest and lowest irreversibilityvalues, respectively. Similar to Fig. 4, the effect ofthe boiling point temperature in this figure doesnot present a consistent trend, as the one showedin the previous figures.The variation of the mass flowrate needed to gen-

erate the same power output with the turbine inletpressure is evaluated in Fig. 8. This figure was gener-ated using the same conditions described in Fig. 6and for a fixed-power output of 50 kW. It can beseen that for all the fluids, the mass flowrateneeded decreases with the increment of the turbineinlet pressure. This is due to the increase in the network of the cycle with the increment of the turbineinlet pressure. These results agree with the resultspresented in Fig. 6, because an increment of thenet work represents an increase in the cycle thermalefficiency. From this figure, it can be observed thatR134a requires the highest mass flowrates amongthe organic working fluids, whereas for pressuresbelow and above 1.5 MPa, R113 and isobutanerequire the lowest mass flowrates, respectively.Figure 9 shows the irreversibility rates for each

fluid for the case analysed in Fig. 8. From thisfigure, it can be seen that irreversibility rates arehigher for low pressures, whereas it decreases withthe increment in the turbine inlet pressure. A slightincrement in the irreversibility rates is observed forhigh turbine inlet pressures. This figure illustratesthat for a pressure range between 0.5 and 3.5 MPa,R113 and R123 show the lowest irreversibility ratesamong the organic fluids.Figure 10 shows the variation of thermal efficiency

with the condenser outlet temperature. This figurewas generated keeping the evaporator pressure con-stant at 3 MPa. The isentropic efficiencies of the

Fig. 7 Variation of the system irreversibility with the

turbine inlet temperature (Tc 298 K) forR113, R123, R245fa, and isobutane

Fig. 8 Mass flowrate needed to produce 50 kW power

versus turbine inlet pressure (Tc 298 K) forR113, R123, R245fa, and isobutane



turbine and pump were 80 and 85 per cent, respect-ively. From this figure, it can be observed that for allthe working fluids, the system thermal efficiencydecreases linearly with the increase in the condenseroutlet temperature. The trend observed in this figureis consistent with the results shown in Figs 3 and 6,where R113 and R123 show the best thermal efficien-cies, whereas propane shows the worst among theevaluated organic fluids. From this figure, the influ-ence of the boiling point on the system thermal effi-ciency can be observed again, as, fluids with higherboiling point temperatures have the best thermalefficiency and vice versa. Figure 11 illustrates thevariation of the system second-law efficiency with

the condenser outlet temperature for R113, iso-butane, and propane. Some fluids were omittedfrom this figure in order to make it clear to read. Itcan be observed that the second-law efficiencydecreases for all the fluids with the increment ofthe condenser outlet temperature. The results pre-sented in Figs 10 and 11 indicate that ORC will bemore beneficial in places with annual low ambienttemperatures, because it will have higher first andsecond law efficiencies.

4 CONCLUSIONS

This article presents an analysis of the performanceof ORC using R134a, R113, R245a, R245fa, R123, iso-butane, and propane and a comparison of theirresults with those of water operating under similarconditions. This analysis was based on the first andsecond laws of thermodynamics, and parameterssuch as thermal efficiency and irreversibility wereevaluated and compared with the results of waterunder the same conditions. It was shown that theexamined organic fluids could be used to generatepower using low-temperature waste heat.Organic fluids need not be superheated as the

cycle thermal efficiency remains approximately con-stant when the inlet temperature of the turbine isincreased. However, using the second-law analysis,it can be seen that superheating organic fluidsincrease the irreversibility. Therefore, organic fluidsmust be operated at saturated conditions to reducethe total irreversibility of the system. It can also beconcluded that the thermal efficiency of ORCincreases when the condenser temperature is

Fig. 9 Total irreversibility rate of the system for a

power output of 50 kW versus turbine inlet

pressure (Tc 298 K) for R113, R123, R245fa,and isobutane

Fig. 10 Variation of the system thermal efficiency

with the condenser outlet temperature

(Pe 3 MPa)

Fig. 11 Variation of the system second-law efficiency

with the condenser outlet temperature

(Pe 3 MPa) for R113, isobutane, and propane



decreased. Therefore, using ORC in locations withlow ambient temperatures will be more effective.Organic dry fluids (R113, R123, R245ca, R245fa, andisobutane) show better performance than wet fluids(R134a and propane). This is because dry fluids donot condense after the fluid goes through the tur-bine, as opposed to wet fluids, which condenseafter the turbine.The influence of the boiling point temperature on

the system thermal efficiency was determined. Thefluid that shows the best thermal efficiency is theone that has the highest boiling point amongthe selected fluids (R113, Tbp 47.59 8C), whereasthe fluid that shows the worst thermal efficiencyhas the lowest boiling point temperature (propane,Tbp 242.09 8C). Therefore, it can be concludedthat the higher the boiling point temperature of theorganic fluid, the better the thermal efficiency thatwill be achieved by the ORC.For the different scenarios analysed in this investi-

gation, ORCs using R113 show the best thermal effi-ciency, whereas those using propane show the worstefficiencies. However, it is important to point outthat some organic fluids show better performancewithin a range of temperatures. Therefore, designershave to closely monitor the operation conditions inorder to select the right organic fluid.

ACKNOWLEDGEMENT

The support from the Department of MechanicalEngineering and the micro-CHP and Biofuel Centerat the Mississippi State University is gratefullyacknowledged.

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APPENDIX

Notation

h specific enthalpy (kJ/kg)_I irreversibility rate (kW)_m mass flowrate (kg/s)q specific heat (kJ/kg)_Q heat rate (kW)s specific entropy (kJ/kg K)S entropy (kJ/K)T temperature (K)TH temperature of the high-temperature

reservoir (K)TL temperature of the low-temperature

reservoir (K)_W power (kW)

DT temperature differential (K)h efficiency (%)

Subscripts

bp boiling pointc condensercycle cyclee evaporatorexit conditions at the exitideal isentropic caseinlet conditions at the inlett turbinep pumpo ambient



iii 21 performance analysis of different working fluids for use in orc

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