Experimental study of the organic rankine cycle under ...faculty.missouri.edu/.../306_Organic_Rankine_Energy... · organic Rankine cycle (ORC) has been largely studied. Due to the

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Energy 180 (2019) 678e688

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Energy

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Experimental study of the organic rankine cycle under different heatand cooling conditions

Hong-Hu Zhang a, Huan Xi a, Ya-Ling He a, *, Yu-Wen Zhang b, Bo Ning a

a Key Laboratory of Thermo-Fluid Science and Engineering of Ministry of Education, School of Energy and Power Engineering, Xi'an Jiaotong University,Xi'an, Shaanxi, 710049, Chinab Department of Mechanical and Aerospace Engineering, University of Missouri, Columbia, MO 65211, USA

a r t i c l e i n f o

Article history:Received 15 March 2019Received in revised form8 May 2019Accepted 11 May 2019Available online 16 May 2019

Keywords:Organic Rankine cycleCycle comparisonWaste heat recoveryEvaporating and condensing temperature

* Corresponding author. Key Laboratory of Thermo-Fof Ministry of Education, School of Energy and PoweUniversity, Xi'an, Shaanxi, 710049, China.

E-mail addresses: [email protected] (H.-H(H. Xi), [email protected] (Y.-L. H(Y.-W. Zhang), [email protected] (B. Ning).

https://doi.org/10.1016/j.energy.2019.05.0720360-5442/© 2019 Elsevier Ltd. All rights reserved.

a b s t r a c t

A small-scaled organic Rankine cycle (ORC) system using R123 as working fluid was experimentallyinvestigated. For ORC and regenerative organic Rankine cycle (RORC), the impacts of the evaporating andcondensing temperatures on the performances of main components (i.e., the expander, the pump, andheat exchangers) and the system were tested. The comparison between these two systems underidentical working conditions was also carried out. The results showed that the expander shaft power ofORC is greater than that of RORC under the identical evaporating temperature and condensing tem-perature. When the evaporating temperature is relatively low, the expander shaft power is more sen-sitive to the condensing temperature, and the thermal efficiency of ORC is higher than that of RORC. Withthe increasing of the evaporating temperature, the thermal efficiency of RORC exceeded that of ORC.Therefore, ORC is recommended for the low temperature heat source, while for the high temperatureheat source, RORC is recommended for its higher thermal efficiency.

© 2019 Elsevier Ltd. All rights reserved.

1. Introduction

Along with the rapid development of science and technology,the consumption of non-renewable energy such as coal, oil, andnature gas increases rapidly, causing serious problems such asenergy shortage and environmental pollution. In order to alleviatethese problems, more attentions are paid to renewable energyutilization, energy conservation, and emission reduction. For high-grade energy utilization, such as nuclear energy and solar energy,supercritical CO2 Brayton cycle has been considered as one of themost promising power cycle because of good stability, compactstructure and small device size [1]. For low-grade energy recovery,organic Rankine cycle (ORC) has been largely studied. Due to thelow boiling point of organics, ORC is suitable to generate electricityby recovering waste heat. In the recent decades, many studies havebeen dedicated to theoretical research of advanced power cycles,

luid Science and Engineeringr Engineering, Xi'an Jiaotong

. Zhang), [email protected]), [email protected]

such as working fluid selection [2e4], performance analysis [5e8],performance studies of critical components [9e11], and combinedcycle analyses [12e14]. There are many researches concerningabout the comparison of ORC and RORC. Li et al. [15] studied theinfluence of IHE and evaporating temperature on ORC using pureand mixture working fluid. They found that the thermal and exergyefficiency had significant improvement after adding an IHE to thesystem. However, the output power decreased when using IHE.Maraver et al. [16] conducted a research to study the effect ofoperating conditions on subcritical and transcritical ORC/RORC.They pointed out that RORC is not applicable if there is no adequatelimitation upon the outlet temperature of heat source. Meinel et al.[17] presented a comparison of two-stage cycle with ORC and RORCutilizing the exhaust gas from an internal combustion engine. Ac-cording to their results, dry fluids is suitable for RORC and isen-tropic fluids is suitable for two-stage cycle. Imran et al. [18] studiedthree different configurations of ORC for geothermal heat source.Their results showed that the basic ORC had the lowest thermal andexergy efficiencies but highest net power and economic perfor-mance. The two-stage ORC had the highest exergy efficiency, andlowest economic performance and net power. According to abovestudies, it is found that there are some disagreement about which isbetter between ORC and RORC. Further experiment research should

Nomenclature

cP specific heat at constant pressure, kJ/(kg$�C)h specific enthalpy, kJ/kgM torque, N$m_m mass flow rate, kg/sn rotational speed, r/min_Q heat transfer capacity, kWT temperature, �C_V volume flow rate, m3/sẆ power, W

Greek symbolsr density, kg/m3

h efficiency

Subscriptcon condensercyc cycleeva evaporatoroil heat transfer oilp pumpt expander

H.-H. Zhang et al. / Energy 180 (2019) 678e688 679

be desired.In order to obtain the practical operating patterns and verify the

model of cycle calculation, many experimental researches havebeen carried out. Lemort et al. [19,20] built and tested an ORCexperimental apparatus. They proposed a method to calculate theparameters of ORC performance. Effects of the internal leakage, thesupply pressure drop and the mechanical losses on expanderisentropic efficiency were compared. They pointed out that internalleakage was the main factor of the expander performance. Li et al.[21] established a RORC system. They found that the thermal effi-ciency of RORC was almost 1.83% higher than that of ORC. Theimpacts of superheat in expander inlet and cooling water flow onthe system performances were also studied [22]. Yang et al. [23]investigated the effects of pressure drop, the degree of super-heating and condensing temperatures on system performance.They found that both the thermal efficiency and the generatingefficiency of the system increased with increasing pressure drop,but condensing temperature exhibited a negative impact on systemperformance. Chang et al. [24] tested an ORC facility using amodified compressor as expander and R245fa as working fluid. Themaximum thermal efficiency and power output were 9.43% and2.3 kW, respectively. Shao et al. [25] investigated the impact ofcondensing parameters on ORC. It was found that the net poweroutput, the pump consumption, and the heat transfer capacities ofevaporator and condenser increased simultaneously withincreasing cooling water flow. Zhou et al. [26] examined an ORC forheat recovery from low-temperature flue gas produced by a liq-uefied petroleum gas stove. The results showed that the systemefficiency, power output, and exergy efficiency increased with theevaporation pressure. However, the increase of superheat willdegrade the system performances.

The expander is the most important part of an ORC. Kang [27]presented a study using a radial turbine and R245fa as the workingfluid. They analyzed the impact of evaporating temperature andexpander pressure ratio on the system performance. Declaye et al.[28] conducted an ORC experiment and evaluated the perfor-mances of expander and system by using a modified open-drive

scroll compressor as an expander. The maximum expander poweroutput, isentropic efficiency, and thermal efficiency were 2.1 kW,75.7%, and 8.5%, respectively. Cho et al. [29] studied the applica-bility of the impulse type turbine. Some nozzles were added at theinlet of the turbine to increase the power output. At the identicalevaporating condition, the power output significantly increased asthe number of nozzles increased. Zhang et al. [30] designed andtested a single-screw expander in ORC to recovery the waste heatfrom a diesel engine exhaust. The results showed that the single-screw expander was more suitable for ORC. Zheng et al. [31]tested an ORC using a rolling-piston expander. It was reported thatthe maximum power output was 0.35 kW, the isentropic efficiencywas 40%, and the system efficiency was between 5% and 6%. Thecomparison of two expanders with different suction volumes arecarried out by our recently work [32]. We found that the expanderwith high suction volume shows a higher shaft power and rotatingspeed. The range of optimal filling factor is 0.8e0.9.

Different working fluids were also adopted in the experimentalresearches. Borsukiewicz-Gozdur [33] designed and built an ORCfacility using R227ea as working fluid, which the electrical effi-ciency could reach 4.88%. R245fa/R365mfc (48.5%/51.5% on a molebasis) was used as the working fluid, and evaluated by Jung et al.[34]. The maximum electrical power and thermal efficiency were0.7 kW and 3.9%, respectively. Qiu et al. [35] established and testeda biomass-fired ORC-based micro-CHP with HFE7000 as theworking fluid. A maximum electrical power of 860.7W, amaximum electrical efficiency of 1.41%, and a maximum CHP effi-ciency of 78.69% were achieved. Desideri et al. [36] compared RORCperformance with two different organics (i.e. SES36 and R245fa).When operating under the identical temperature difference be-tween evaporator and condenser, the system using R245fa as theworking fluid could generate more power than the one usingSES36. The working fluids, expanders and working conditions inthe above studies are summarized in Table 1. It can be found fromTable 1 that R123 had been applied in many experimental re-searches [19,21,25,26]. Some theoretical studies also reported thatbetter performance could be obtained by using R123 as theworkingfluid. Maizza et al. [37] calculated the thermodynamic efficienciesat different evaporating and condensing temperatures. They foundthat R123 and R124a showgood system performance under varyingoperating conditions among the fluids they analyzed. Roy et al. [38]analyzed an ORC utilizing the waste heat of flue gas at 140 �C and312 kg/s. The results showed that ORC using R123 as working fluidcould obtain the maximum power and both first and second lawefficiency among all the selected fluids. Wang et al. [39] investi-gated the effects of some parameters, such as heat source tem-perature, evaporating and condensing pressures, on the systemperformance under the optimal conditions. They suggested R123 asworking fluid for ORC at the temperature ranges from 100 �C to180 �C because of its shorter payback period. Therefore, R123 wasused as the working fluid in this study.

From the above literature review, it can be found that althoughthere are numerous experimental studies of ORC, most of themwere carried out under certain working conditions. Some of themconsidered the impact of heat source or evaporating temperatureon system performance. A few papers discussed the comparison ofORC and RORC. Little information can be found on experimentalresearch on the impact of condensing temperature coupled withthe evaporating temperature on ORC and RORC. In this paper, anORC experimental system using R123 as the working fluid wasbuilt, and the system can be switched between ORC and RORCmodes. The expander adopted in this platform is scroll expander,which was modified from a scroll compressor. In the section ofResults and Discussions, the impacts of the evaporating andcondensing temperatures with respect to the performances of the

Table 1List of parameters of the experimental investigations.

Author(s) Working fluid Type of expander Temperature range

Lemort et al. [19,20] R123 Open-drive scroll expander Expender inlet: 101.7e165.2 �CLi et al. [21,22] R123 Impulse axial flow type turbine with single stage Heat source: 85, 95, 110, 130 �C

Condensing: 25 �CYang et al. [23] R245fa Open-drive scroll expander Heat source: 100 �C

Condensing: 21.86e43.63 �CChang et al. [24] R245fa Scroll expander Heat source: 77.9 �C

Cooling water: 11.5e21 �CShao et al. [25] R123 Radial inflow turbine Heat source: 72 �C

Condenser outlet: 26e51 �CZhou et al. [26] R123 Scroll expander Heat source: 90e220 �CKang [27] R245fa Radial turbine Evaporation: 77.1e82.3 �C

Condensing: 37.4e40.3 �CCho et al. [29] R245fa Impulse turbine Expender inlet: 50e115 �CZheng et al. [31] R245fa Rolling-piston expander Heat source: 90 �C

Condensing: 23 �CBorsukiewicz-Gozdur [33] R227ea Hermetic turbogenerator Heat source: 66.8e110 �CJung et al. [34] R245fa/R365mfc (48.5%/51.5%) Scroll expander Heat source: 100e150 �C

Condensing: 18e20 �CQiu et al. [35] HFE7000 Vane-type air motor Heat source: 117.8e128.9 �CDesideri et al. [36] R245fa, SES36 Single screw expander Heat source: 125 �C

Cooling glycol water: 14e43 �C

H.-H. Zhang et al. / Energy 180 (2019) 678e688680

main equipment, including the expander, the pump and the heatexchangers would be firstly investigated. Secondly, the variations ofthe systemperformancewith different condensing and evaporatingtemperature are presented. In addition, the comparisons of thecharacteristics between ORC and RORC are also discussed.

2. ORC/RORC experimental system

Fig. 1 shows the schematic diagram of ORC/RORC experimentalsystem. The system contains three subsystems: (a) ORC/RORCsubsystem, (b) the heat source subsystem, and (c) the cooling watercirculation subsystem. The picture of the system is shown in Fig. 2.The ORC/RORC subsystem consists of an evaporator, an expander, aregenerator, a condenser, a pump, two liquid storage tanks, severalvalves, and pipes. The scroll compressor using in automobile aircondition system is modified as an expander because of its char-acteristics of low cost [40], compact structure, few moving parts,

Fig. 1. Schematic diagram of ORC

and usability in the gas-liquid two-phase region. The brazed plateheat exchangers are chosen due to its high heat transfer coefficientand compact structure. The pump adopted in this test system is adiaphragm-metering pump that the flow rate can be controlled byregulating the piston stroke. In the evaporator, the organic workingfluid becomes high-pressure vapour by absorbing heat from heat-transfer oil (HTO). It then expands in an expander to generate po-wer. In the regenerator, heat exchanging takes place between theworking fluid from the outlet of the pump and expander. After that,it is cooled down in the condenser, pressurized in the pump, andreturns to the evaporator. In this system, ORC or RORC can beswitched by adjusting Valves 3 and 4. The main part of the heatsource subsystem is an electrical heater. The HTO is heated in theheater and then is pumped into the evaporator to simulate a low-temperature heat source. The cooling water circulation subsystemconsists of a water pump, a glass rotameter, a cooling tower, and awater storage tank.

/RORC experimental system.

Fig. 2. Photograph of ORC/RORC.

Table 2Operating conditions of the experiment system.

Parameters Range

Evaporating temperature 80e110 �CHTO volume flow rate 7.5m3/hCondensing temperature 23e30 �CTorque-load of dynamometer 0e6 Nm

H.-H. Zhang et al. / Energy 180 (2019) 678e688 681

3. Parameters of the experiment

3.1. Operating conditions and theoretical analysis

Evaporating temperature (ET, temperature at the outlet of theevaporator) and condensing temperature (temperature at theoutlet of the condenser) are the main experimental parameters in

(a) ORC (b

Fig. 3. T-s diagram of ORC and RORC w

this study. The ET is controlled by adjusting the power of the heater,while the condensing temperature is adjusted through changingthe flow rate of the cooling water. The measurement is conductedwhen the system is working under steady state. Each steady state ismaintained at least 10min. Table 2 shows the main operatingconditions of the experiment.

Fig. 3 shows the T-s diagrams of the experimental system in ORCand RORC mode. Eight measuring points are set up in this system.For points 1 to 6, the temperature and pressure are measured, butfor point 7 and 8, only the temperature is measured. REFPROP 9.0[41] is employed to calculate thermodynamic properties such astemperature, pressure, enthalpy, entropy, and density.

The system performances are quantified by the followingparameters:

Evaporator heat transfer capacity:

) RORC

ith superheating and subcooling.

Table 3Parameters of measuring equipment.

Name Range Accuracy

Type-T thermocouple �200-350 �C ±0.5 �CPressure sensor 0e2.5MPa ±0.2% Full scaleGlass rotor flowmeter 0.4e4m3/h ±2.5% Full scaleVortex shedding flowmeter 0e10m3/h ±1.0% Full scaleDynamometer and secondary instrument rotation speed: 0e9999 rpm

torque: 0e10 Nmrotation speed: ±0.2% Full scaletorque: ±1.0% Full scale

H.-H. Zhang et al. / Energy 180 (2019) 678e688682

_Qeva ¼ _moilcpoilðT7 � T8Þ ¼ _VoilroilcpoilðT7 � T8Þ (1)

The mass flow rate of working fluid:

_m ¼ _Qeva

.ðh1 � h6Þ (2)

Expander shaft power:

_Wt ¼p$M$n=30 (3)

Condenser heat transfer capacity:

0 1 2 3 40

50

100

150

200

250

300

Shaftp

ower

(W)

Torque (N·m)

Condensingtemperature

23°C24°C25°C26°C27°C28°C29°C30°C

(a) ET: 80°C

0 1 2 3 4 5 6 7200

250

300

350

400

450

500

550

Shaftp

ower(W

)

Torque (N·m)

Condensingtemperature

23°C24°C25°C26°C27°C28°C29°C30°C

(c) ET: 100°C (d

Fig. 4. Effect of evaporating and condensing

_Qcon ¼ _mðh3 � h4Þ (4)

Pump power:

_Wp ¼ _mðh5 � h4Þ (5)

Pump isentropic efficiency:

hs ¼ ðh5 � h4Þ=ðh5 � h4Þ (6)

System thermal efficiency:

hcyc ¼�

_Wt � _Wp

�._Qeva (7)

0 1 2 3 450

100

150

200

250

300

350

400

450

Shaftpow

er(W

)

Torque (N·m)

Condensingtemperature

23°C24°C25°C26°C27°C28°C29°C30°C

(b) ET: 90°C

0 1 2 3 4 5 6350

400

450

500

550

600

650

Shaftpow

er(W

)

Torque (N·m)

Condensingtemperature

24°C25°C26°C27°C28°C29°C30°C

) ET: 110°C

temperature at different torque for ORC.

0 1 2 350

100

150

200

250Shaftpow

er(W

)

Torque (N·m)

Condensingtemperature

23°C24°C25°C26°C27°C28°C29°C30°C

1 2 3 4 50

100

200

300

400

Condensingtemperature

23°C24°C25°C26°C27°C28°C29°C30°C

Shaftpow

er(W

)

Torque (N·m)

(a) ET: 80°C (b) ET: 90°C

0 1 2 3 4 5150

200

250

300

350

400

450

500Condensingtemperature

23°C24°C25°C26°C27°C28°C29°C30°CSh

aftpow

er(W

)

Torque (N·m)0 1 2 3 4 5 6

300

350

400

450

500

550

600

Shaftpow

er(W

)

Torque (N·m)

Condensingtemperature

24°C25°C26°C27°C28°C29°C30°C

(c) ET: 100°C (d) ET: 110°C

Fig. 5. Effect of evaporating and condensing temperature at different torque for RORC.

22 24 26 28 30 32 342.2

2.4

2.6

2.8

3.0

3.2

3.4

Turbine

pressureratio

Condensing temperature (°C)

ET(ORC)80°C90°C100°C110°C

ET(RORC)80°C90°C100°C110°C

Fig. 6. Expander pressure ratio versus condensing temperature at different ET.

22 23 24 25 26 27 28 29 30 31 32 33 340.060

0.065

0.070

0.075

0.080

0.085

0.090

Working

fluidflowrate(kg/s)

Condensing temperature (°C)

ET(ORC)80°C90°C100°C110°C

ET(RORC)90°C100°C110°C

Fig. 7. Variations of the working fluid flow rate with condensing temperature atdifferent ET.

H.-H. Zhang et al. / Energy 180 (2019) 678e688 683

22 23 24 25 26 27 28 29 30 31

0

3

6

9

12

15ET for ORC for RORC

80°C 80°C90°C 90°C100°C 100°C110°C 110°C

Superheatin

gdegree(°C)

Condensing temperature(°C)

Fig. 8. Superheating degrees at expander inlet under different condensing tempera-tures and ETs.

H.-H. Zhang et al. / Energy 180 (2019) 678e688684

According to equation (1), _Qeva can be calculated by the specificheat at constant pressure (cp,oil), the volumetric flowrate ( _Voil), thedensity (roil) of the heat transfer oil, T7 and T8. Where, the cp,oil androil can be calculated by the average temperature of T7 and T8 ac-cording to the follow equations:

cp;oil ¼ � 4:0732� 10�6T2ave;oil þ 7:3� 10�3Tave;oil � 0:059776

(8)

roil ¼ � 0:49589Tave;oil þ 1009:2 (9)

Tave;oil ¼ ðT7 þ T8Þ=2 (10)

It can be found that the _Qeva is only dependent on T7, T8 and _Voil,in which the temperatures at states 7 and 8 are measuredaccordingly.

22 23 24 25 26 27 28 29 30 310.11

0.12

0.13

0.14

0.15

0.16

Pumpinletpressure(M

Pa)

Condensing temperature (°C)

Evaporating temperature (ET)80°C 90°C 100°C 110°C

Solid: ORCShot dash: RORC

(a) inlet pressure Fig. 9. Variations of pump pressure

3.2. Instrumentation and uncertainty analysis

Type-T thermocouples and pressure sensors were adopted tomeasure temperature and pressure, respectively. A dynamometer isemployed to measure the rotation speed and torque. A vortexshedding flowmeter and a glass rotor flowmeter are assembled inthe heat source subsystem and the cooling water circulation sub-system, respectively. Table 3 shows the parameters of the mea-surement devices.

Based on the previous researches [23,24] and the error propa-gation theory, the root-sum-square method is carried out tocalculate the measuring uncertainties:

DF ¼ffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiXi

�vFvxi

Dxi

�2vuut (11)

where F denotes the parameters calculated by the measuring pa-rameters xi, DF and Dxi denote the uncertainty of F and xi,respectively.

4. Results and discussions

4.1. Expander performances

Figs. 4 and 5 present the variation of the expander shaft powerat different shaft torques for ORC and RORC, respectively. It can beseen that when the shaft torque increases, the shaft power ofexpander first increases, achieves a maximum value, then de-creases. This trend indicates that there exists an optimal shafttorque corresponding with the highest shaft power. The shaft po-wer of ORC is higher than RORC. The maximum shaft power of ORCand RORC are 611.2W and 577.6W, respectively. Fig. 4(a) and (d)show the variations of shaft power with ET of 80 �C and 110 �C,respectively. When ET is fixed at 80 �C, raising the condensingtemperature from 23 �C to 30 �C leads themaximum shaft power todecrease from 261.0W to 133.9W, which reduces approximately48.7%. However, when ORC is operated at an ET of 110 �C, the shaft

22 23 24 25 26 27 28 29 30 310.4

0.5

0.6

0.7

0.8

0.9

1.0

1.1

1.2

Pumpoutletpressure(M

Pa)

Condensing temperature (°C)

Evaporating temperature (ET)80°C 90°C 100°C 110°C

Solid: ORCShot dash: RORC

(b) outlet pressurewith condensing temperature.

22 23 24 25 26 27 28 29 30 317

8

9

10

11

12

ET for ORC for RORC80°C 80°C90°C 90°C100°C 100°C110°C 110°C

Subcooling(°C)

Condensing temperature (°C)

Fig. 11. Variations of subcooling at pump inlet with condensing temperature atdifferent ETs.

H.-H. Zhang et al. / Energy 180 (2019) 678e688 685

power reduces only 3.6% as condensing temperature increases from23 �C to 30 �C. This illustrates that the condensing temperature hasa great impact on the shaft power under lower ET. The effect ofcondensing temperature on the shaft power decreases with theincrease of ET. When ET increases from 80 �C to 100 �C for ORC, thesensitivity of shaft power to condensing temperature decreasesrapidly. The relative variation of maximum shaft power withdifferent condensing temperatures is less than 5% when ET isgreater than or equal to 100 �C. However, for RORC, the relativevariation is 17.4% when ET is 100 �C; moreover, the relative varia-tion is less than 5% until ET is 110 �C. In conclusion, it is necessary tofocus on the effect of condensing temperature on the expanderperformance, especially at low evaporating temperature. The upperlimit of the evaporating temperature for RORC should be higherthan that of ORC, when studying the influence of condensingtemperature on expander performance of these two systems.

Fig. 6 illustrates the effect of condensing temperature on theexpander pressure ratio under different ETs. It is clearly observedthat the expander pressure ratio decreases with increasingcondensing temperature, and increases with increasing ET for bothsystems. The pressure ratio of ORC is greater than that of RORC atthe same ET. The back pressure of the expander for ORC is lowerthan RORC because of the regenerator. It is the main reason causingthe reduction of the expander pressure ratio. The reduction of theexpander pressure ratio is one of the main reasons that caused thedecrease of the expander shaft power.

Fig. 7 shows the effect of condensing temperature on the massflow rate of working fluid at different ETs. It is obvious that themassflow rate of RORC is higher than that of ORC. With increasingcondensing temperature, the pump outlet pressure and theexpander inlet pressures increase. As a result, the density of theworking fluid at the entrance of the expander increases. Therefore,the mass flow rate of the working fluid increases correspondingly.Some test results are not marked in Fig. 7, especially the results at ahigh condensing temperature and low evaporating temperature.The reason for this is that theworking fluid goes into the two-phasestate at the condition described above, which means that thethermodynamic parameters of working fluid could not be deter-mined by temperature and pressure. Therefore, the specificenthalphy of the working fluid is uncertain, and the mass flowcould not be calculated.

22 23 24 25 26 27 28 29 30 3110

15

20

25

30

35

40

45

50ET of RORC system

80°C 90°C 100°C 110°CET of ORC system

80°C 90°C 100°C 110°C

Pumpisentropicefficiency(%

)

Condensing temperature (°C)

(a) isentropic efficiency Fig. 10. Variations of isentropic efficiency and pump p

Fig. 8 illustrates the superheating degree at expander inlet withdifferent ETs and condensing temperatures. It is clearly observedthat the superheating degree at the inlet of the expander increaseswith increasing of ET and decreases with increasing of condensingtemperature for both systems. The increase of the condensingtemperature causes the condensing pressure to rise, and theevaporating pressure increases correspondingly. Therefore, thesuperheating degree at expander inlet reduces. By comparing theresults of ORC and RORC, it can be seen that under the sameworking conditions, the superheating at the inlet of the expanderfor RORC is lower, which leads the working fluid easier to enter thetwo-phase region. When the working fluid enters the two-phaseregion, which is drawn in the figure with superheating degree of0 �C, it is difficult for working fluid to fully expand in the expander,resulting in the deterioration of the heat-work conversion capacityof the expander, which is the main reason for the decline of itsoutput power.

22 23 24 25 26 27 28 29 30 3180

100

120

140

160

180

200ET of ORC system

80°C 90°C 100°C 110°CET of RORC system

90°C 100°C 110°C

Pumpconsum

ption(W

)

Condensing temperature (°C)

(b) pump power consumptionower consumption with condensing temperature.

23 24 25 26 27 28 29 30 31180

210

240

270

300

Evaporator enthalpy difference

Evaporator outlet specific enthalpy

Evaporator intlet specific enthalpy

Condensing temperature (°C)

ORCRORC

Evaporatorinletspecific

enthalpy

(kJ/kg) E

vaporatoroutletspecific

enthalpy(kJ/kg)

Evaporator

enthalpydifference

( kJ/kg)

430

435

440

445

450

180

210

240

270

300

Fig. 13. Variations of evaporator specific enthalpy and enthalpy difference withcondensing temperature.

H.-H. Zhang et al. / Energy 180 (2019) 678e688686

4.2. Pump performances

Fig. 9 shows the variations of inlet and outlet pressures of thepump with condensing temperature for ORC and RORC. The inletpressure of the pump of ORC is higher than that of RORC, but thepump outlet pressure of ORC is lower than that of RORC. In RORC,the regenerator acts as a “pre-condenser”, which makes the overallcooling heat exchange capacity of RORC greater than that of ORC.The condensing pressure at condenser outlet for RORC are thereforeslightly lower. Meanwhile, the flow resistance of the system in-creases due to the addition of a regenerator in RORC. In order toensure the stability of the system, the pressure at the pump outletincreases accordingly.

Fig. 10 shows the power consumption and isentropic efficiencyof the pump in ORC and RORC as a function of evaporating andcondensing temperature. As the condensing temperature increases,the isentropic efficiency of the working fluid pump in both ORC andRORC increase. Moreover, the increasing rate of pump isentropicefficiency decreases when the evaporating temperature changesfrom 80 �C to 110 �C, as the solid (for ORC) or the hollow (for RORC)symbols shown in Fig. 10 (a). It can also be found that the pumpisentropic efficiency of ORC is higher than that of RORC. As shown inFig. 10 (b), the pump power consumption decreases with theincreasing of the condensing temperature. The reduction of thepump power consumptionwith condensing temperature decreaseswhen the evaporating temperature increases from 80 �C to 110 �C.Nevertheless, when evaporating temperature is 110 �C, the powerconsumption of the working fluid pump shows a fluctuating vari-ation. It is because when the evaporating temperature is at a lowlevel, the isentropic efficiency of the pump increases with theobvious increase of the condensing temperature. The isentropicefficiency is the main factor in influencing the pumping power, sothe pump power consumption shows a trend of decline. However,when the evaporating temperature increases, the increment of thepump isentropic efficiency reduces. Meanwhile, the mass flow rateof the working fluid increases with increasing condensing tem-perature. The higher isentropic efficiency gives a lower specificenthalpy difference and hence a less pumping power consumption,but the greater mass flow causes a higher pump consumption. The

Fig. 12. Rupture of pump diaphragm by cavitation.

combined effect of the pump isentropic efficiency and the massflow rate makes a fluctuation in the pumping power.

Fig. 11 shows the variations of subcooling at pump inlet withevaporating and condensing temperatures. The subcooling at pumpinlet decreases slightly with the increase of the condensing tem-perature, and higher ETs could correspond to greater subcooling.Furthermore, it can be found that the subcooling of ORC is higherthan that of RORC. It is because that RORC has lower condensingpressure, as shown in Fig. 9, its saturated liquid temperature andthe subcooling is also lower as a result. The range of subcooling inthis work is about 7e10.5 �C. According to the results in Ref. [42],the recommended subcooling should be greater than 20 �C to avoidcavitation. It is difficult to obtain such a large subcooling due tocondensing conditions in this work. Therefore, the cavitation oc-curs in the pump, which was the main factor causing the low ef-ficiency of the working fluid pump. Pressure fluctuation caused bycavitation during the experiment had also caused damage to thediaphragm of pump, as shown in Fig. 12.

4.3. Heat exchangers performances

Fig. 13 shows the variations of inlet/outlet specific enthalpiesand enthalpy differences of the evaporator with condensing tem-perature in ORC and RORC when the ET is 110 �C. It can be seen thatfor ORC, inlet specific enthalpy of the evaporator increases with thecondensing temperature, but the specific enthalpy at evaporator

23 24 25 26 27 28 29 30 31330

360

390

420

450

Condenser enthalpy difference

Condenser outlet specific enthalpy

Condenser intlet specific enthalpy

Condensing temperature (°C)

ORCRORC

Condenserinletspecific

enthalpy

(kJ/kg) C

ondenseroutletspecific

enthalpy(kJ/kg)

Condenser

enthalpydifference

(kJ/kg)

210

220

230

240

250

175

205

235

265

295

Fig. 14. Variations of condenser specific enthalpy and enthalpy difference withcondensing temperature.

22 23 24 25 26 27 28 29 30 31 32 33 341.0

1.5

2.0

2.5

3.0

Therm

alefficiency(%

)

Condensing temperature(°C)

ET(ORC)80°C90°C100°C110°C

ET(RORC)90°C100°C110°C

Fig. 15. Variations of the thermal efficiency with condensing temperature at differentETs.

H.-H. Zhang et al. / Energy 180 (2019) 678e688 687

outlet has a slighter variation than the inlet. Therefore, the enthalpydifference decreases with increasing condensing temperature. ForRORC, due to the buffer effect of the regenerator, the variation ofcondensing temperature has less influence on the inlet specificenthalpy of the evaporator. The outlet specific enthalpy of theevaporator decreases due to the increase of the outlet pressure. Theenthalpy difference of the evaporator decreases consequently.Comparing the results of ORC with that of RORC, the evaporatorenthalpy difference of RORC is significantly lower than that of ORC.The presence of the regenerator makes the evaporator inlet specificenthalpy of RORC higher than that of ORC. Meanwhile, the outletpressure of the evaporator in RORC is higher than that of ORC. Theevaporator outlet specific enthalpy of RORC is lower as a result.

Fig. 14 illustrates the variation of inlet/outlet specific enthalpiesand enthalpy differences of the condenser with condensing tem-perature in ORC and RORC when the ET is 110 �C. The condenserinlet specific enthalpy shows little variationwith the increase of thecondensing temperature, but the specific enthalpy of the outlet hasincreased significantly, which makes the enthalpy difference of thecondenser exhibited a trend of decrease. Due to the effect of theregenerator, the condenser inlet specific enthalpy of RORC issignificantly lower than that of ORC. Moreover, the outlet pressureof the condenser as well as the specific enthalpy of the workingfluid for both systems are relatively close at the same condensingtemperature. These are the reasons why the condenser enthalpydifference in RORC is significantly lower than that in ORC.

4.4. System performances

Fig. 15 indicates the effect of condensing temperature on thethermal efficiency of the two systems at different ETs. The thermal

Table 4System performances at partial operating conditions.

Conditions Systems Evaporator capacity/W Sha

ET: 90 �CCondensing temperature: 24 �C

ORC 13345 375RORC 12499 325

ET: 110 �CCondensing temperature: 24 �C

ORC 16571 590RORC 15228 578

ET: 90 �CCondensing temperature: 25 �C

ORC 13930 383RORC 12793 315

ET: 110 �CCondensing temperature: 25 �C

ORC 16557 594RORC 15187 572

efficiency increases with increasing ET for both systems. Throughthe comparison of the experiment results of both systems, it can befound that the thermal efficiency of ORC is slightly higher than thatof RORC when the ET is lower than 90 �C. The thermal efficiency ofRORC gradually exceeds that of ORC with the increase of the ET.

Table 4 shows the performances of the two systems, when theevaporating temperatures are 90 �C and 110 �C, the condensingtemperatures are 24 �C and 25 �C, respectively. The heat transfercapacity of the evaporator of ORC is greater than that of RORC underthe same operating condition, due to the impact of the regenerator.The expander shaft power of ORC is also greater than that of RORCunder the same condition. It can also be found that, under the samecondensing temperature, the relative variation of the expandershaft power of the two systems at lower ET are greater than that athigher ET. When the ET is 90 �C, the relative variation of theexpander shaft power is greater than that of the heat transfer ca-pacity of the evaporator. Therefore, the expander shaft power is themajor factor of the system performances. When the ET is 110 �C, therelative variation of the expander shaft power is lower than that ofthe evaporator's heat transfer capacity. The heat transfer capacity ofthe evaporator becomes the major factor of the system perfor-mances for this case, and the thermal efficiency of RORC exceedsthat of ORC.

5. Conclusions

In this study, the performances of an ORC/RORC to utilize thelow-temperature waste heat is experimentally studied. A modifiedscroll compressor is chosen as the expander of the system, andR123 is used as the working fluid. The system performances underdifferent heat and cooling conditions are investigated. The mainconclusions are as follows:

(1) The expander shaft power of ORC is greater than that of RORCunder the fixed evaporating temperature and condensingtemperature.Within the range of temperature studied in thispaper, the maximum shaft power of RORC and ORC are611.2W and 577.6W, respectively. The effect of thecondensing temperature on the performance reduces withthe increase of evaporating temperature. For ORC, thecondensing temperature has less impact on the expandershaft power (with a relative variation < 5%) when theevaporating temperature is greater than or equal to 100 �C.However, for RORC, the evaporating temperature should begreater than or equal to 110 �C when the impact ofcondensing temperature is insignificant.

(2) Under the given working conditions in this work, the massflow rate of RORC is higher than that of ORC. Thermal effi-ciencies of both systems increase with the increasing ofevaporating temperature. When the evaporating tempera-ture is 90 �C, the thermal efficiency of ORC is slightly higher

ft power/W Pump power/W Net power/W Thermal efficiency/%

144 231 1.73124 201 1.61161 429 2.59147 431 2.83133 250 1.7993 222 1.73168 426 2.57122 450 2.97

H.-H. Zhang et al. / Energy 180 (2019) 678e688688

than that of RORC. With a higher evaporating temperature,the thermal efficiency of RORC would exceed that of ORC.

(3) Adding a regenerator in the organic Rankine cycle may notimprove the thermal efficiency. When the evaporating tem-perature is low, the regenerator would reduce the pressureratio of the expander, thus reducing the output power andthe thermal efficiency of the system. However, the influenceof the regenerator on the power output of the expander isreduced when the evaporating temperature is high, and thethermal efficiency of RORC is higher than that of ORC.Therefore, it is recommended to adopt ORC when the heatsource temperature is low but RORC when the heat sourcetemperature is high to ensure the high thermal efficiency ofthe system.

Acknowledgements

The work was supported by the Key Project of National NaturalScience Foundation of China (No.51436007).

The authors would also like to thank the Foundation for Inno-vative Research Groups of the National Natural Science Foundationof China (No.51721004) and Science and Technology PlanningProject of Xi'an (201809160CX1JC2-02).

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