Applied Energy 87 (2010) 3366–3373

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Applied Energy

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A comparative study of pure and zeotropic mixtures in low-temperature solarRankine cycle

J.L. Wang, L. Zhao *, X.D. WangDepartment of Thermal Energy and Refrigeration Engineering, School of Mechanical Engineering, Tianjin University, No. 92 Weijin Road, Tianjin 300072, PR China

a r t i c l e i n f o a b s t r a c t

Article history:Received 22 March 2010Received in revised form 10 May 2010Accepted 12 May 2010Available online 17 June 2010

Keywords:Zeotropic mixtureRankine cycleSolarLow temperature

0306-2619/$ - see front matter Crown Copyright � 2doi:10.1016/j.apenergy.2010.05.016

* Corresponding author. Tel.: +86 22 81590264; faxE-mail address: jons@tju.edu.cn (L. Zhao).

The paper presents an on site experimental study of a low-temperature solar Rankine cycle system forpower generation. The cycle performances of pure fluid M1 (R245fa) and zeotropic mixtures M2

(R245fa/R152a, 0.9/0.1) and M3 (R245fa/R152a, 0.7/0.3) are compared, respectively, based on the exper-imental prototype. The experiments have been conducted under constant volume flow rate of differentfluids. The results show that, with the component of R152a increasing, the system pressure levelincreases and the power output varies accordingly, which provides an additional means of capacityadjustment. The collector efficiency and thermal efficiency of zeotropic mixtures are comparativelyhigher than pure fluid of R245fa in the experimental condition, which indicates that zeotropic mixtureshave the potential for overall efficiency improvement. Due to the non-isothermal condensation of zeo-tropic mixture, the condensing heat could be partially recovered by adding an external heat exchanger.Thus, compared with pure fluid R245fa the system overall efficiency of zeotropic mixtures could beimproved.

Crown Copyright � 2010 Published by Elsevier Ltd. All rights reserved.

1. Introduction

In the energy-to-power conversion industry, the thermal effi-ciency becomes uneconomically low when the gaseous steam tem-perature drops below 700 F (371 �C), the bulk of the energy losesand systems using water as working fluid become significantly lessefficient and capital cost increase [1]. To overcome this disadvan-tage and still exploit the basic Rankine cycle technology developedover the years, systems based on working fluids such as hydrocar-bons or refrigerants are being developed and researched in recentyears. This is so called organic Rankine cycles (ORCs). The organicfluids is most promising for Rankine cycle power generation sys-tems in that they can use low grade heat from a variety origins,such as geothermal energy [2,3], solar radiation [4], biomass com-bustion [5], waste heat from the industrial process [6–8]. It shouldbe pointed out that, in the solar powered Rankine cycle researcharea, some important solar Rankine cycle systems have been devel-oped and examined recent years, such as using freon R245fa [9]and natural fluid-CO2 [10–13] as working fluid.

In the design of the cycle system, one important factor shouldbe considered is the working fluid selection. Regarding the work-ing fluids, several studies envisage the utilization of differentworking fluids that are more appropriate for low temperatureheat source than tradition steam power cycles, such as, benzene

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[14], n-pentane [15], isobutane [16], and some cryogens like thefreons R11, R113, R114, Rl23 and R134a [17–20].

It can be seen that the previous studies of ORCs regarded moreabout single component organic fluids. However, an importantlimitation of pure fluids is the constant temperature evaporationwhich is not suitable for sensible heat sources such as waste heat.The mixtures have variable temperature during the phase changeprocess, which could be used to reduce the mismatch of tempera-ture profiles between heat transfer fluid and the evaporating orcondensing working fluid mixtures. Thus, the system irreversibili-ties can be minimized. Another advantage of the mixture is that itcould be acquired the fluids that have the same thermodynamicproperties as the pure ones through different component mass fric-tion. This could greatly extend the range of candidate working fluidselection for low temperature Rankine cycle.

Presently, it can be noticed that there is a significantly arisinginterest in multi-component mixture research. Radermacher dem-onstrated the mutual influence of working fluid mixtures proper-ties on Rankine cycle performance, and counter-flow heatexchangers are suggested in the system for the mixtures [21].Angelino and Paliano compared n-pentane and mixture of n-bu-tane and n-hexane (50%/50%) by simulating liquid geothermal re-source for electricity generation. The results show that mixtureyields 6.8% more electricity than n-pentane and 25% less air areused with potential benefits in both cooler frontal area and fanpower consumption [22]. Aleksandra studied Rankine cycle withheat source temperature of 80–115 �C by using different working

ights reserved.

Nomenclature

A collector aperture area, m2

h1, h02 enthalpy of different state point in T–S diagram, kJ/kgI global solar radiation, W/m2

_m mass flow rate of working fluids, kg/s_QE heat absorbed in the collector, kW_WP power consumed by working fluid pump, kW_WT power generated by expander, kW

gC collector efficiency, %gR Rankine cycle efficiency, %gT turbine efficiency, %gsys overall efficiency, %

J.L. Wang et al. / Applied Energy 87 (2010) 3366–3373 3367

fluids, both natural and synthetic as well as mixtures [23]. Wangand Zhao made a theoretical analysis of zeotropic mixturesR245fa/R152a used in low-temperature solar Rankine cycles [24].

However, since the research in this field is in a very early stage,very few experimental studies on low temperature Rankine cyclefor power generation using zeotropic mixtures are reported. In or-der to analyze the thermal performance of low-temperature solarRankine cycle system using zeotropic mixtures, the paper concen-trates on relative experimental researches. An experimental proto-type is constructed and tested, such as the one that this paperdescribed. The flat-plate collector, which does not need a complextracking system and thus has low cost, has been considered, andthe experimental evaluation and cycle performances could be atool for their further optimization.

Fig. 1. T–S diagram of different working fluids.

2. Working fluid selection

In ORCs applications, the choice of working fluid is importantsince the fluid must not only possess thermophysical propertiesthat match the application but also have adequate chemical stabil-ity at the desired working temperature. Here, R245fa and R152aare chosen as the components of the mixture because they bothhave zero ODP (Ozone Depression Potential) and lower GWP (Glo-bal Warming Potential), which have less environmental impact.The property details of the compositions are shown in Table 1.Here, pure fluid R245fa is noted as M1 and the mass fraction of zeo-tropic mixtures M2 and M3 are 0.9/0.1 and 0.7/0.3, respectively. Un-der this mass fraction, the fluids have comparatively low criticalpressure, which is acceptable in the experimental condition. Thetemperature–entropy diagrams of these fluids are shown inFig. 1. In Fig. 1, pressures of 2.0 MPa and 0.2 Mpa for different fluidsare drawn, respectively. It is clear that as the mass fraction ofR152a increases the pressure level rises and at the same time itis desirable to have the high pressure resistant equipment. The sys-tem economic efficiency needs to be further revalued.

Fig. 2. Basic solar Rankine cycle for electricity generation.

3. System description

3.1. Experimental apparatuses

Fig. 2 illustrates the basic solar Rankine cycle system. The ther-modynamic process can be briefly described as follows: the liquidorganic working fluid is compressed with a working fluid pumpwhich forces the fluid through the flat-plate collector. Heat trans-fers from solar radiation to the liquid, and the fluid is preheated,

Table 1Property details of working fluids.

Fluid Molecular weight (g/mol) Normal boiling pointa (�C) Critical pre

R245fa 134.05 15.14 3.65R152a 66.05 �24.02 4.52

a The normal boiling point is boiling point at a pressure of 1 atmosphere.

vaporizes and finally superheats at the collector outlet. High pres-sure vapor is directed to the expander which is coupled with a gen-erator. The exhaust vapor of the expander passes through thecondenser. Condensation and a small degree of sub-cooling occur.Liquid is then sucked by the pump and the cycle recommences.

ssure (MPa) Critical temperature (�C) ODP GWP Safety group

154.01 0 950 B1113.26 0 120 A2

Fig. 3. Experimental layout of the low-temperature solar Rankine cycle.

3368 J.L. Wang et al. / Applied Energy 87 (2010) 3366–3373

The experimental prototype mainly consists of two flat plate so-lar collector, a throttling valve, an air-cooled condenser, a storagetank, a fluid pump, a radiometer and date acquisition system.Fig. 3 shows the layout of the experimental system.

Because the power output of the experimental prototype is toosmall, there is no real expander available for this system. There-fore, a throttling valve is used in the system to expand the vaporinstead of a real expander. By adjusting the valve opening state,the pressure difference of the realistic expander can be simulatedaccordingly. Obviously, the fluid temperature and its thermody-namic and transport properties and states at the valve outletare different from those of the true expander condition. However,the cycle performance for the power production can be derivedbased on thermodynamic analysis. As the fluid passing throughthe valve is a throttling process, the power generated from theexpander can be calculated by setting the expander efficiency as0.85 [25].

During the experiment, only one flat-plate collector, with itsinternal tube tank diameter of 12 mm, is used. The collector aper-ture area is 0.6 m2. It possesses one glass cover sheet and a highefficiency solar absorber plate whose absorptivity is 0.95 andemissivity is 0.17. Fig. 4 presents a photo of the experimentalprototype.

Fig. 4. A photo of the experimental prototype.

3.2. Data acquisition system

The system is instrumented comprehensively. Each set of mea-surements is taken in a time step of 1 min and the logged data arerecorded as a function of time. The testing parameters include tem-perature, pressure, flow rate and solar radiation.

� Temperature T-type thermocouples with accuracy of ±0.5 �C aremounted in the system to provide temperature measurement.The testing points are shown in Fig. 3 and the ambient temper-ature is also recorded.� Pressure transducers provide pressure values for collector

outlet and throttling outlet. The accuracy of the transducerof points 1 and 2, as shown in Fig. 3, is ±3 kPa and ±1 kPa,respectively.� Solar radiation to measure the global solar radiation, a pyra-

nometer with accuracy less than ±0.05 W is installed on the45� slop surface which has the same angle of collector.

The output signals of these instruments are connected to a PCthrough an Agilent 34980A series data logger and data are recordedas functions of time simultaneously. A diaphragm metering pump,which could be regulated in the range of 0–100% of rated flow, is usedin the prototype to feed the fluids and control volume flow rate.

4. Assumption and calculation method

The paper presented is the results of calculations regarding theRankine cycle, which is shown in Fig. 3. Thermodynamic parame-ters of working fluids have been taken from the computer programof REFPROP8.0 [26], which is developed by the National Institute ofStandards and Technology.

Some additional parameters must be set to perform the calcula-tion in this case:

1. Each component is considered as a steady-state-flow system.2. The kinetic and potential energies as well as friction losses are

neglected.3. The expander isentropic efficiency gT is set at 0.85.4. Furthermore, composition shift is neglected in the experimental

system.5. No pressures drop in the heat exchangers and pipelines have

been considered.

For the conditions, the cycle T–S diagram of pure fluid R245faand zeotropic mixtures are illustrated in Figs. 5 and 6, respectively.

Fig. 5. Schematic diagram for R245fa Rankine cycle.

Fig. 6. Schematic diagram for zeotropic mixtures Rankine cycle.

J.L. Wang et al. / Applied Energy 87 (2010) 3366–3373 3369

Based on the above measured parameters, it is possible to get aset of calculated values for pure fluid R245fa and zeotropic mix-tures, and thus obtain a global understanding of the system behav-ior and performance. These values include:

Power generated by the expander _WT ¼ gT � _m � ðh1 � h20 Þ ð1Þ

Power consumed by the working fluid pump _WP

¼ _m � ðh5 � h40 Þ ð2Þ

Heat absorbed in the collector _Q E ¼ _m � ðh1 � h5Þ ð3Þ

Collector efficiency gC ¼_QE

I � A ð4Þ

Rankine cycle efficiency gR ¼_WT � _WP

_QE

ð5Þ

Overall efficiency gsys ¼ gC � gR ¼_WT � _WP

I � A ð6Þ

where h1, h20 , h40 and h5 are the enthalpies of corresponding statepoints in T–S diagram. _m is the fluid mass flow rate of the system,which could be converted from the experimental volume flow rate.And, I and A represent the global solar radiation on the collectors’inclination surface and collector aperture area, respectively. Itshould be pointed out that the overall efficiency gsys in Eq. (6) isthe ratio of the net power output to the solar radiation on the col-lector’s surface.

In order to evaluate the cycle performances of different workingfluid, the volume flow rate was set as constant of 1.3 L/h and the

Fig. 7. Solar radiation during the experiment.

valve open state was also kept in the same open state withoutbeing adjusted.

5. Results and discussion

The experimental procedure was conducted in spring (April2009) in the city of Tianjin, China. Fig. 7 shows the solar radiationfor the 1st, 6th and 16th of April 2009, respectively. It can be seenthat the solar radiation of 1st and 6th are almost identical, and themaximum value are both 1055 W/m2 which appeared at11:48 a.m. and 12:02 a.m., respectively. Affected by the air quality,the overall solar radiation of 16th is slightly lower and the maxi-mum value is appeared at 12:02 a.m. with value of 1001 W/m2.Compared with 1st and 6th, the solar radiation of 16th does not re-sult to an extreme change of system operation and this radiationdeviation is almost insignificant, which could be neglected.

Fig. 8 shows the collector inlet and outlet temperature of differ-ent working fluids during the experiment. Since the process ofphase change occurs in the collector, the collector outlet tempera-ture is greatly influenced by solar radiation and ambient tempera-ture. Thus, the outlet temperatures fluctuate during theexperiment. However, under the daily constant flow rate, the over-all trend of the outlet temperatures of different fluids follows thatof the solar radiation, and the maximum outlet temperatures are105.9 �C (M1, 12:56 a.m.), 99.86 �C (M2, 12:55 a.m.) and 101.56 �C(M3, 13:03 p.m.), respectively. During the experiment, an interest-ing observation concerning the behavior of the system is that com-pared with the time of maximum value of solar radiation, ahysteresis of the fluids maximum outlet temperature is observed.It could be noticed that the time lag is about 1 h. The identifiedtime lag of hysteresis is directly related to the system thermal iner-tia of solar collector. The same phenomenon could be found in thefollowing operation in the afternoon, that is, when the solar radia-tion descends from the maximum, collector outlet temperature ex-hibit no trend of decreasing but maintain the same states of hightemperature. It is because that the collector internal tube tankand thermal insulation material act as a thermal storage duringthe experiment and heat is stored when the solar radiation is suf-ficient. Though there is collector heat loss, the working fluids couldstill get the sufficient heat for fluids evaporating and superheatingand the collector outlet temperatures could achieve high values.When the stored heat and solar radiation are lower than collectorheat loss, the outlet temperatures begin to drop, so as the degree ofsuperheating. Therefore, the time of outlet temperature beginningto decrease is also behind that of solar radiation by 1 h. Accordingto Manolakos [4], the thermal inertia of the system makes theoperation more stable and smoothly reacted to the variation of so-lar radiation.

Fig. 8. Collector inlet and outlet temperature.

Fig. 10. Ambient temperature and condenser inlet and outlet temperature.

3370 J.L. Wang et al. / Applied Energy 87 (2010) 3366–3373

In Fig. 8, as the collector inlet temperature of M1 is compara-tively higher during the experiment, the outlet temperature ofM1 is higher than M2 under the approximately similar solar radia-tion. Though the solar radiation of 16th April has small deviationcompared with 1st and 6th, the outlet temperature of M1 still reachhigh value.

Fig. 9 shows the pressure of different working fluid in the exper-imental monitoring point. As mentioned above, the pressure levelof the mixtures increases with the adding of R152a. Obviously,M3 has the highest pressure level and M1 has the lowest. Influencedby the collector outlet temperature fluctuation, the collector outletpressures of three fluids show the similar variation. It should benoted that, for zeotropic mixtures M2 and M3, the compositionshift, which probably leads to composition change during thephase change process, results in collector outlet pressure fluctuat-ing acutely during the operation. The mechanism of this phenom-enon needs further study. Since the air-cooled condenser is used inthe prototype, the pressures after the valve have the relationship toambient temperature and thus behave smooth trend.

The ambient temperature, together with the condenser inletand outlet temperature, is illustrated in Fig. 10. Condenser inlettemperature, namely the valve outlet temperature, has the greatrelationship to the collector outlet temperature as the fluid passingthe valve is a process of throttling. At the same time, affected bythe ambient temperature the condenser outlet temperatures in-crease with the ambient temperature ascending. The condenseroutlet temperatures of three working fluids are all higher thanambient temperature about 2 or 3 �C on average. The ambient tem-perature of Fig. 10a is higher than that of Fig. 10b and c, whichleads to a higher collector inlet temperature of M1 in Fig. 8.

Fig. 11 shows different degree of superheating at collector out-let. It can be seen that the collector outlet states of M1 are super-heating in the whole experimental process. However, the outletstates of M2 and M3 could be divided into three stages: liquid–va-por phase at the beginning, superheating in the steady state, li-quid–vapor phase at the end of the experiment. This fact couldbe explained that, compared with M2 and M3, M1 has the highercollector inlet temperature and lower latent heat during evapora-tion, which brings on the large degree of superheating. As theworking fluid phase change process plays an important role inthe whole cycle, when incident radiation is lower, heat absorptionin the collector will decrease and therefore boiling is incomplete.This will have a negative effect on the system operation, whichcould be seen in the following analysis.

In order to illustrate the superheating of different fluid undersame constant flow rate, Table 2 presents the degree of superheat-ing of different fluids during the experiment. Pure working fluid M1

Fig. 9. Collector outlet pressure and after valve pressure.

has the maximum superheating of 54.89 �C which is higher thanM2 and M3 of 7.06 �C and 13.87 �C, respectively. Accordingly, M1

has the maximum average degree of superheating. This result incollector heat loss increases and collector efficiency decreases. Be-cause a high degree of superheating in the expander inlet wouldnot improve Rankine cycle efficiency [8], the Rankine cycle effi-ciency of M1 does not show any special improvement. It deservesto mention that an optimized regulation of fluid flow rate is re-quired in order to minimize the superheating heat loss, which inturn increases the collector efficiency.

The heat absorbed by the working fluids in the collector couldbe calculated from Eq. (3). Enthalpy difference, together with heatabsorption in the flat-plate collector, is illustrated in Fig. 12. Duringthe period of 8:34–9:24 (6th April) and 8:34–9:41 (16th April), asthe fluid do not finish the phase change process in the collector,fluid M2 and M3 have the lower heat absorption, so as the enthalpydifference. From Fig. 12a the calculation results show that, in the

Fig. 11. Superheating at the collector outlet.

Table 2Degree of superheating during the experiment.

Working fluid M1 M2 M3

Time of superheating 8:39–15:22

9:50–15:40

9:40–15:19

Maximum degree of superheating(�C)

54.89 47.83 41.02

Average degree of superheating (�C) 35.19 27.86 26.37

Fig. 12. Specific enthalpy difference and heat absorbed by the fluid in the collector.

Fig. 13. Collector efficiency.

J.L. Wang et al. / Applied Energy 87 (2010) 3366–3373 3371

superheating period of 9:50–15:19, the averaged specific enthalpydifference of M3 is 281.17 kJ/kg, which is higher than M2’s258.31 kJ/kg and M1’s 248.97 kJ/kg. Therefore, under the similar so-lar radiation the zeotropic mixtures have large specific heatabsorption. Fig. 12b shows the actual heat absorption of differentworking fluid under the same volume flow rate of 1.3 L/h. It can

be seen that fluids heat absorption are almost identical duringthe superheating period. This can be explained that, with thecomponent of R152a increasing, the mixtures have lower criticaldensity and the mass flow rates are different correspondingly.According to the calculation, under the volume flow rate of 1.3 L/h the mass flow rate of M1, M2 and M3 is 1.73 kg/s, 1.69 kg/s and1.51 kg/s, respectively. Therefore, if the fluids were in the samemass flow rate, zeotropic mixtures will show a potential to im-prove the collector efficiency.

Fig. 13 shows the collector efficiency during the experiment. Asmentioned above, in the periods of 8:34–9:24 (6th April) and 8:34–9:41 (16th April), M2 and M3 are in the liquid–vapor states and thecollector efficiencies are lower as 4.19% and 12.27%, respectively.From 9:50 to 15:19, the fluids are all in the state of superheating

3372 J.L. Wang et al. / Applied Energy 87 (2010) 3366–3373

and the collector efficiency curves exhibit the shape of concave.The minimum point of the concave curve is the time of maximumsolar radiation. It’s because that, when the solar radiation reachesits maximum value at noon, the heat loss of collector also reachesits maximal, and thus the collector efficiency shows the minimalvalue. When the solar radiation begin to fall, just as the systemthermal inertia mentioned above, the attenuation of the collectoroutlet temperature lags behind that of the solar radiation and thecollector efficiencies behave the trend of ascending.

Integrating Figs. 7 and 12b, Fig. 13 shows that the collector effi-ciencies of three fluids are almost the same in the period of super-heating. M3 shows slightly higher collector efficiency than M1 andM2 by 7.91% and 6.45%, respectively, as is presented in Table 3.

In order to evaluate the cycle performance of different workingfluid, it deserves to concentrate on the states of superheating from9:50 to 15:19, which are the steady states of three working fluids.The following analyses are based on this period.

The power produced from the expander is shown in the follow-ing Fig. 14. As zeotropic mixture M3 has the lower superheating de-gree at collector outlet, a comparatively higher power output isobtained at the maximum of 9.06 W. Affected by the fluctuationof collector outlet temperature and pressure, the power output ofM2 varies from 4.69 W to 7.69 W. The average values could befound in Table 3. The average power output of M3 is higher thanthat of M1 and M2 by 29.10% and 28.03%, respectively. It can beseen that zeotropic mixtures have the great ability for power out-put, and the system capacity adjustment could be easily realizedunder different composition.

The Rankine cycle efficiencies and the overall efficiencies can becalculated from Eqs. (5) and (6). Table 3 summarizes a set of char-acteristic value of cycle performance for different working fluids,

Fig. 14. Power output from the expander.

Table 3Comparison of different fluid cycle performance.

Working fluid M1 (April1st)

M2 (April6th)

M3 (April16th)

Period of superheating 9:50–15:19

9:50–15:19

9:50–15:19

Average heat absorbed in thecollector, W

118.87 120.17 118.86

Average power consumed by workingfluid pump, W

1.04 0.88 1.07

Average power output, W 5.98 6.03 7.72Average collector efficiency, % 21.25 21.54 22.93Average Rankine cycle efficiency, % 4.16 4.29 5.59Average overall efficiency, % 0.88 0.92 1.28

which is useful to get an idea of the overall performance of the sys-tem. The averaged Rankine cycle efficiency of M3 is the highest of5.59%. On the contrary, M1 has the lowest of 4.16%. The systemoverall efficiency has a great relationship to collector efficiencyand Rankine cycle efficiency, so the average overall efficiency ofM1, M2 and M3 is 0.88%, 0.92% and 1.28%, respectively. These lowvalues, in some extent, are in close relation to the lack of regulationof flow rate of working fluids. As the zeotropic mixtures have thelarge amount of condensation heat, measures could be taken to re-cover part of condensation heat. For example, an external heat ex-changer could be added between the expander and condenser torecover low temperature heat which could be used for domestichot water supply. Due to the temperature glide of zeotropic mix-tures during condensation, they would get the hot water whoseoutlet temperatures are higher than that of pure fluids. Therefore,the overall efficiency of zeotropic mixtures can be improved.

6. Conclusion

An investigation on low-temperature solar Rankine cycle per-formance has been performed based on an experimental prototypeby using zeotropic mixtures and pure fluid R245fa. Based on theconstant flow rate of the fluid, the current work mentioned aboveis largely dedicated to make a system comparison between fluidsM1, M2 and M3 to identify the cycle performance. The main resultsof the experimental measurements can be extracted as follow:

� The thermal inertia of the collector leads to the maximum out-let temperature lagged behind that of the solar radiation about1 h.� Under sufficient solar radiation to evaporate the working fluids

in the collector, the flow rate is an important factor affecting thecycle performance. In the following procedure, a further optimi-zation of flow rate regulation is needed.� Due to the larger latent heat of evaporation, fluid M3 has large

specific enthalpy difference in the collector, which has a greatpotential to improve the collector efficiency comparing withthe pure R245fa.� In the experimental superheating period, the average power

output of M3 is higher than that of M1 and M2 by 29.10% and28.03%, respectively. It can be seen that the power output variesaccordingly and the system capacity adjustment could be easilyrealized under different composition of zeotropic mixtures.� The average overall efficiency of M1, M2 and M3 is 0.88%, 0.92%

and 1.28%, respectively. For zeotropic mixtures, there is a greatpotential to improve the overall efficiency by introducing anexternal heat exchanger to recover the partial condensationheat.

Acknowledgement

The authors would like to acknowledge the financial supportprovided by the Program for New Century Excellent Talents inUniversity.

Appendix A. Uncertainty analysis

In every measured parameter there is an error between themeasured and real value. Therefore, in order to evaluate the exper-iment data thoroughly and more reliable, an uncertainty analysesis necessary. In the following Table A there are the measuredinstruments and their uncertainty. Table B shows the total experi-mental uncertainty for every calculated value derived from mea-sured data.

J.L. Wang et al. / Applied Energy 87 (2010) 3366–3373 3373

Table A. Measured parameters and uncertainty.

Instruments

Parameters Range Uncertainty(%)

Pyranometer

Solar radiation 0–2000(W/m2)

5.0

Working fluidpump

Volume flow rate

0–6.5 (L/h)

5.0

Thermocouples Collector inlet 10– 1.5

temperature 100 �C Collector outlettemperature

10–100 �C

2.0

Condenser inlettemperature

10–100 �C

2.0

Condenser outlettemperature

10–100 �C

1.5

Pump inlettemperature

10–100 �C

1.5

Pump outlettemperature

10–100 �C

1.5

Pressuretransducer

Collector outletpressure

0–1.5 MPa

1.0

Throttling outletpressure

0–0.5 MPa

0.7

Table B. Uncertainty of every variable.

Calculated variable

Total experimentaluncertainty (%)

Power generated by the expander

5.72 Power consumed by the working

fluid pump

5.56

Heat absorbed in the collector

5.68 Collector efficiency 7.57 Rankine cycle efficiency 6.59 Overall efficiency 8.27

References

[1] Marciniak TJ, Krazinski JL, Bratis JC, Bushby HM, Buycot EH. Comparison ofRankine-cycle power systems: effects of seven working fluids. ArgonneNational Laboratory Report ANL/CNSV-TM-87, Illinois; 1981.

[2] Umberto D, Gianni B. Study of possible optimization criteria for geothermalpower plants. Energy Convers Manage 1997;38:1681–91.

[3] Madhawa Hettiarachchi HD, Mihajlo G, William MW, Yasuyuki I. Optimumdesign criteria for an organic Rankine cycle using low-temperature geothermalheat sources. Energy 2007;32:1698–706.

[4] Manolakos D, Kosmadakis G, Kyritsis S, Papadakis G. On site experimentalevaluation of a low-temperature solar organic Rankine cycle system for ROdesalination. Solar Energy 2009;83:646–56.

[5] Drescher U, Brüggemann D. Fluid selection for the organic Rankine cycle (ORC)in biomass power and heat plants. Appl Therm Eng 2007;27:223–8.

[6] Wei DH, Lu XS, Lu Z, Gu JM. Performance analysis and optimization of organicRankine cycle (ORC) for waste heat recovery. Energy Convers Manage2007;48:1113–9.

[7] Liu BT, Chien KH, Wang CC. Effect of working fluids on organic Rankine cyclefor waste heat recovery. Energy 2004;29:1207–17.

[8] Meng Xiangyu, Yang Fusheng, Bao Zewei, Deng Jianqiang, Serge Nyallang N,Zhang Zaoxiao. Theoretical study of a novel solar trigeneration system basedon metal hydrides. Appl Energy 2010;87(6):2050–61.

[9] Wang XD, Zhao L, Wang JL, Zhang WZ, Zhao XZ, Wu W. Performance evaluationof a low-temperature solar Rankine cycle system utilizing R245fa. Solar Energy2010;84:353–64.

[10] Zhang XR, Yamaguchi H, Uneno D. Experimental study on the performance ofsolar Rankine system using supercritical CO2. Renew Energy 2007;32:2617–28.

[11] Zhang XR, Yamaguchi H, Uneno D, Fujima K, Enomoto M, Sawada N. Analysis ofa novel solar energy powered Rankine cycle for combined power and heatgeneration using supercritical carbon dioxide. Renew Energy 2006;31:1839–54.

[12] Wang Jiangfeng, Sun Zhixin, Dai Yiping, Ma Shaolin. Parametric optimizationdesign for supercritical CO2 power cycle using genetic algorithm and artificialneural network. Appl Energy 2010;87(4):1317–24.

[13] Cayer Emmanuel, Galanis Nicolas, Desilets Martin, Nesreddine Hakim, RoyPhilippe. Analysis of a carbon dioxide transcritical power cycle using a lowtemperature source. Appl Energy 2009;86(7–8):1055–63.

[14] Hung TC, Shai TY, Wang SK. A review of organic Rankine cycles (ORCs) for therecovery of low-grade waste heat. Energy 1997;22:661–7.

[15] Nguyen VM, Doherty PS, Riffat SB. Development of a prototype low-temperature Rankine cycle electricity generation system. Appl Therm Eng2001;27:169–81.

[16] Pedro JM, Louay MC, Kalyan S, Chandramohan S. An examination ofregenerative organic Rankine cycles using dry fluids. Appl Therm Eng2008;28:998–1007.

[17] Bertrand FT, George P, Gregory L, Antonios F. Fluid selection for a low-temperature solar organic Rankine cycle. Appl Therm Eng 2009;29:2468–76.

[18] Tchanche BF, Lambrinos Gr, Frangoudakis A, Papadakis G. Exergy analysis ofmicro-organic Rankine power cycles for a small scale solar driven reverseosmosis desalination system. Appl Energy 2010;87(4):1295–306.

[19] Quoilin Sylvain, Lemort Vincent, Lebrun Jean. Experimental study andmodeling of an organic Rankine cycle using scroll expander. Appl Energy2010;87(4):1260–8.

[20] Chacartegui R, Sanchez D, Munoz JM, Sanchez T. Alternative ORC bottomingcycles FOR combined cycle power plants. Appl Energy 2009;86(10):2162–70.

[21] Radermacher R. Thermodynamic and heat transfer implications of workingfluid mixtures in Rankine cycles. Int J Heat Fluid Flow 1989;10:90–102.

[22] Angelino G, Paliano PCD. Multicomponent working fluids for organic Rankinecycles (ORCs). Energy 1998;23:449–63.

[23] Aleksandra B et al. Comparative analysis of natural and synthetic refrigerantsin application to low temperature Clausius–Rankine cycle. Energy2007;32:344–52.

[24] Wang XD, Zhao L. Analysis of zeotropic mixtures used in low-temperaturesolar Rankine cycles for power generation. Solar Energy 2009;83:605–13.

[25] Tutrboden. High efficiency Rankine for renewable energy and heat recovery.<http://www.turboden.it/en/orc.asp> [accessed 28.04.09].

[26] Lemmon EW, Huber ML, McLinden MO. Reference fluid thermodynamic andtransport properties-REFPROP. Standard Reference Datebase 23, Version 8.0,National Institute of Standard and Technology; 2007.