Experimental verification of a rolling-piston expander that applied for low-temperature Organic Rankine Cycle

Applied Energy 112 (2013) 1265–1274

Contents lists available at SciVerse ScienceDirect

Applied Energy journal homepage: www.elsevier.com/locate/apenergy

Experimental verification of a rolling-piston expander that applied for low-temperature Organic Rankine Cycle N. Zheng, L. Zhao ⇑, X.D. Wang, Y.T. Tan Key Laboratory of Efficient Utilization of Low and Medium Grade Energy, MOE, Tianjin University, Tianjin 300072, China

h i g h l i g h t s " A new type of kW-sized expander for low-temperature ORC system is proposed. " A low-temperature ORC system with heat source temperature of 90 °C is set up. " A dynamic test of the rolling-piston expander is conducted. " Influence factors on the performance of the expander are investigated. " A maximum output shaft power of 0.34 kW and a steady expander isentropic efficiency of 43.3% are obtained.

a r t i c l e

i n f o

Article history: Received 22 September 2012 Received in revised form 6 December 2012 Accepted 11 December 2012 Available online 5 January 2013 Keywords: Rolling-piston expander Low temperature Organic Rankine Cycle R245fa Experiment

a b s t r a c t A kilowatt-sized rolling-piston expander is proposed for low-temperature Organic Rankine Cycle power generation system. The main purpose of this paper is to conduct a dynamic test of the proposed expander. A low-temperature ORC system using R245fa as working fluid is established for the performance evaluation of the proposed expander. In the experimental system, a hot-water boiler is applied to supply a low temperature heat source. A variety of working conditions are achieved by varying the feed pump capacity and the generator load. In these tests, experimental data such as the R245fa flow rate, the temperature of expander inlet and outlet, the pressure of expander inlet and outlet, the output power and rotation speed of the expander are measured and analyzed. The experimental results show that the proposed expander normally runs between 350 rpm and 800 rpm with a maximum output power of 0.35 kW when the heat source temperature is below 90 °C. Meanwhile, a maximum expander isentropic efficiency of 43.3% and a stable Rankine cycle efficiency of 5% are also obtained. Ó 2012 Elsevier Ltd. All rights reserved.

1. Introduction Because of the potential in reducing the consumption of fossil fuels and alleviating the environmental problems, utilizing renewable energy and industrial waste heat to generate electricity has attracted much attention in the past decade. To date, the ORC-based power generation is considered to be the most promising technology [1] and has been widely applied to the power generation from low temperature heat sources, such as solar energy [2], geothermal energy [3,29], biomass energy [4], industrial waste heat [5,30] and exhaust gas of internal combustion engine [6,31]. When considering a low temperature ORC-based power generation system, a certain challenge is to select an applicable expander, because the expander is a critical component limiting the Rankine cycle efficiency. The selection of expander should be depended on the operating conditions and the magnitude of the ⇑ Corresponding author. Tel.: +86 022 27404188. E-mail address: [email protected] (L. Zhao). 0306-2619/$ - see front matter Ó 2012 Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.apenergy.2012.12.030

power output. Since the available low-temperature heat sources that ORCs convert in mechanical energy are usually small, expanders with output power below 10 kWe are considered. The ORC-based power generation system with the size in the range of 1–10 kWe can be applied to a Combined Heat and Power (CHP) system to supply heat and power demands of typical domestic buildings and light commercial buildings. In comparison with the fossil fuel-powered micro-scale CHP systems such as the Stirling enginebased CHP systems, the internal combustion engine-based CHP system, and the PEM fuel cell-based CHP system, the ORC-based CHP systems are more attractive because they are not powered by fossil fuels and hence can save more primary energy and CO2 emissions [25]. A number of theoretical and experimental studies on the microscale expander that applied in ORC-based power generation system have been reported in the past decades. Lemort et al. developed a scroll expander semi-empirical model to compute variables of first importance and pointed out that the internal leakages, the supply pressure drop and the mechanical losses were the main

1266

N. Zheng et al. / Applied Energy 112 (2013) 1265–1274

Nomenclature Symbols A h N n Pe V

gm gR gs q

sectional area, m2 enthalpy, kJ/kg torque, N m rotation speed, rpm shaft work, kWe volume, m3 expander mechanical efficiency, % Rankine cycle efficiency, % expander isentropic efficiency, % density, kg/m3

losses affecting the performance of the expander [7]. Quoilin et al. performed a numerical and experimental study on an ORC which used R123 as working fluid. The expander mounted in their experimental system was originally an oil-free open-drive scroll compressor, adapted to operate in reverse [8]. Qiu et al. presented preliminary experimental investigations of a biomass-fired microCHP with ORC. A modified air motor as expander was mounted in their system and an electric efficiency of 1.34% was achieved [9]. In another paper, the findings of the market research for expanders for micro-CHP systems with ORC were summarized, and the working principles and characteristics of different types of expanders were introduced and evaluated [10]. Stefano Clemente et al. presented a detailed model of a scroll machine, and a series of comparisons were carried out to evaluate how real scroll expander characteristic affect the expected efficiency and power output of the ORC system [26]. Daniele Fiaschi et al. implemented a step by step design model of a radial turbo expander for ORC considering the real fluid properties. Hence, the influences of the selected working fluid on the cycle thermodynamics and the turboexpander efficiency could be reflected [27]. Pei et al. constructed a kW-scale ORC system with a specially manufactured turbine using R123 as working fluid. During their dynamic tests, a turbine isentropic efficiency of 65% and a cycle efficiency of 6.8% were obtained. In addition, several problem-solving techniques were presented, such as the avoidance of cavitations in the pump [11]. Wang et al. proposed a concept which combined an ORC with a conventional vapor compression cycle. A 5 kW cooling capacity prototype system with a modified scroll compressor operating as expander using R245fa was developed and tested [12]. Although different types of expanders offer merits in certain respects and may be selected, micro-scale expanders (<10 kWe) are not commercially available at present. Hence, researchers often turn to those specifically in-house designed and manufactured machines or modified compressors. However, only a handful of experimental studies of low-temperature ORC equipped with such an expander could generate electricity. A list of recent experimental works on micro-scale expanders using ORC is shown in Table 1. In general, scroll expanders show relatively high efficiencies (usually from 48% to 65% [8]) when applied in small-scale power generation system with a power range of 1–10 kWe. Scroll expanders are normally redesigned based on available scroll compressors. Limited by the structure and size of the scroll unit, scroll expanders have rarely been applied in power generation systems with a power range of 20–50 kWe. Besides, a complex lubrication system has to be designed when operating scroll compressors in expansion mode [13]. Turbine expanders are generally applied in power cycles with power output greater than 50 kWe [32] and their efficiencies are likely to become unacceptable when scaled down to 10 kWe level. In addition, smaller turbine expanders are very

h Va Vt

rotor angle, ° actual expander displacement, m3/s theoretical expander displacement, m3/s

Subscripts EXP expander EVP evaporator I inlet O outlet

expensive and operate at very high rotation speed where reliability has yet to be proven [28]. The screw expanders are also good choices for large and medium scale power cycles and they are more cost-effective than turbines. However, due to the difficulty in sealing the organic working fluid, the performance of screw expanders deteriorates markedly as they are scaled down to 20 kWe level. Since the micro-scale expanders are currently under development, the authors hope to develop a brand-new type of expander applied in ORC-based power generation systems. The rolling-piston expander proposed by the authors could address some of the drawbacks encountered with the existing expanders. The rolling-piston expander has simple structure and comprises mainly cylindrical components, which simplifies the manufacture process and hence reduces the cost. Besides, rolling-piston expanders run at lower rotation speed than turbines and could operate normally without an auxiliary lubrication system. What is more, the rolling-piston expander could tolerate high inlet pressure (9 MPa). Hence, this type of expander could be applied in power generation systems with higher power output, which indicates the application area of rolling-piston expander is wider than that of scroll expander. A rolling-piston expander was applied for the first time in a low-temperature solar Rankine cycle for power generation, and the experimental results indicated its appropriateness to microscale ORC-based power generation system [2]. However, dynamic test of this type of expander has not been conducted by far. For the purpose of the present study, the performance of the R245fa rolling-piston expander under varying conditions is verified experimentally based on a low-temperature ORC system. Parameters such as rotation speed, output shaft power and isentropic efficiency of the expander, and the overall ORC efficiency are evaluated. All these recorded data will be used to improve the expander performance by modifying the geometrical characteristic of the expander and optimizing the system layout in the next work. 2. The proposed rolling-piston expander The rolling-piston expander was originally designed and manufactured for small-sized refrigeration system and heat pump system by Zha [14], and it was first applied in a trans-critical carbon dioxide refrigeration cycle system by Yang et al. [15]. Fig. 1 shows the photos of the rolling-piston expander which is specially designed and produced for a low-temperature ORC using R245fa as working fluid. The expander seems to be bulky, because a starter motor is assembled together with the expander. Fig. 2 shows the basic structure of the R245fa rolling-piston expander. The expander is composed of these main parts: a cylinder, a rolling-piston, an eccentric rotor, a sliding vane and an inlet control valve. As shown in the figure, the crescent-shaped cavity between the cylinder and the rolling-piston is the working chamber. The

1267

N. Zheng et al. / Applied Energy 112 (2013) 1265–1274 Table 1 List of recent experimental work on micro-scale expander using ORC. Participant

Working fluid

Temperature of heat source (°C)

Type of expander

Volume (pressure) ratio

Shaft power (kWe)

Rotation speed (rpm)

Expander efficiency (%)

Saitou [21] Qiu [4] Quoilin [8] Peterson [22] Pei [11] Kane [23] Wang [2] Yamamoto [24]

R113 HFE-7000 R123 R123 R123 R123/R134a R245fa R123

140 >127 101–163 183 105 150 110 70

Scroll Vane Scroll Scroll Vane Scroll Rolling piston Turbine

– 2.1 2.7–5.4 4.6 7.2 2.3 4 3

0.45 1.72 0.4–1.8 0.256 1 9 1.64 0.15

1800 852 1771–2660 1400 60,000 – 900 35,000

63 55.45 42–68 50 65 67 45 –

Fig. 2. Basic structure of the rolling-piston expander. 1. Shell, 2. Sealing ring, 3. Cylinder, 4. Discharge port, 5. Inlet port, 6. Sliding vane, 7. Rolling piston, 8. Eccentric rotor, 9. Control valve. Table 2 Specifications of the rolling-piston expander.

Fig. 1. Photos of the rolling-piston expander: (a) appearance; (b) main parts breakdown illustration.

sliding vane separates the working chamber into two parts: the high-pressure chamber which connects the inlet port and the low-pressure chamber which connects the outlet port. The inlet control valve will shut after a certain amount of working fluid enters the high-pressure chamber. The precise moment that the valve shuts is determined by the designed volume ratio of the expander. The principal specifications of the rolling-piston expander are listed in Table 2. To describe the working process of the expander clearly, the rotor angle (h) is introduced, as shown in Fig. 3. The piston starts to

Structural parameter

Symbol

Value

Unit

Cylinder inside diameter Piston outside radius Cylinder axial length Sliding vane thickness Eccentricity

D r H B e

152 61 76 13 15

mm mm mm mm mm

move from its top position where the rotor angle is set to 0°. Then the piston rolls along the cylinder clockwise. The moment the rotor angle reaches 28°, the inlet control valve opens and the high temperature, high pressure vapor R245fa begins to enter the working chamber. The suction process continues until the rotor angle reaches 170°. During the suction process, the vapor R245fa expands and converts the internal energy into mechanical energy by driving the piston to roll. Finally, the low temperature, low pressure vapor is exhausted through the expander outlet port. Actually, the suction process, the expansion process and the exhaust process are accomplished simultaneously, which contributes to the low flow velocity at the inlet and outlet of the expander. As a result, the expander operates smoothly. For a low rotation speed expander, leakage loss is the main factor that affects its performance. For the purpose of smooth running, appropriate clearances should be allowed between the kinematic pairs. However, these clearances may lead to the leakage of the working fluid. The main leaking locations inside the expander are also marked in Fig. 3 (marked as M1, M2 and M3). To minimize the leakage loss, some measures are taken. For example, to diminish the internal leakage M1, high pressure vapor is guided to the

1268

N. Zheng et al. / Applied Energy 112 (2013) 1265–1274

Fig. 3. The geometrical relationship of the working chamber and the main leakage location.

back-end of the sliding vane to maintain an appropriate clearance between the piston and the sliding vane. The lubricating oil films formed between the piston and the cylinder will reduce the internal leakage M3. Besides, o-type sealing rings were applied to avoid the external leakage of the working fluid, as shown in Fig. 2. The sealing rings were made of EPDM rubber (Ethylene Propylene Diene Monomer rubber) and showed good properties in terms of strength and corrosion resistance. Fig. 5. (a) Gas tank, liquid tank, condenser and evaporator. (b) R245fa feed pump and expander.

3. Experimental apparatus and method To conduct dynamic tests of the rolling-piston expander, a low-temperature ORC system using R245fa as working fluid is designed and constructed. Fig. 4 is the schematic diagram of the

P T 8 2

T

11

T 1

P T

T

6

10

T T

5 T

T

T 4

9 M

T 3

12 T

Fig. 4. Schematic diagram of the experimental system.

7

experimental system. The experimental system mainly comprises of an evaporator, an electric hot-water boiler, a feed pump, a liquid tank, a condenser, a cooling tower, two water pumps, a rollingpiston expander and the interrelated measurement and data acquisition system. Fig. 5 shows the photos of the prototype system. An appropriate heat source is a matter of great concern for micro-scale ORC-based power generation systems. In theory, different types of low-temperature heat sources such as solar energy, geothermal energy and biomass energy could be utilized by the prototype system. Taking the aspects such as initial and operation cost, the limitation on CO2 emissions, and the government policies into consideration, we think the solar energy is the most promising and appropriate heat source for such system. In this study, however, a hot-water boiler instead of solar collectors is used as the low-temperature heat source to eliminate the influence of the weather on the test. The hot-water boiler possesses a rated power output of 36 kW and can provide hot water from 40 °C to 90 °C. The diaphragm metering pump used in the experimental system to feed liquid R245fa is supplied by AILIPU (model JYMZ 630/1.3). The pump can provide a maximum operating pressure of 1.3 MPa, flow rate of 630 L/h and rotation speed of 120 rpm. The maximum power output of the pump is 1.5 kW, and the output capacity of the pump can be adjusted and recorded in the range of 0–100%. Plate heat exchangers supplied by ALFA LAVAL (model

1269

N. Zheng et al. / Applied Energy 112 (2013) 1265–1274

CB76-H) are used in the heating and cooling processes. A mechanical draft water cooling tower (cooling capacity 100 kW) is used as heat sink of the experimental system, dissipating waste heat to the environment. During the experiment, the rolling-piston expander is coupled to a high-speed DC generator, while some incandescent bulbs are used as load. Standard T-type thermocouples, with an accuracy of ±0.2 °C, are used to measure R245fa temperatures at different positions in the prototype. The temperature measuring points are shown in Fig. 4. Two pressure transmitters with an accuracy of ±0.2% are installed to measure pressures. A GPGXE6B1NM1 type transmitter with measurement range of 0–1.5 MPa is used to measure the expander inlet pressure and a GPGXE5B1NM1 type transmitter with measurement range of 0–0.5 MPa is used to measure the expander outlet pressure. In order to measure the rotation speed and output shaft power of the expander, a torque transducer supplied by CATIC Beijing Electrical Wister (model CYB-803S) is installed. The measurement range of torque and rotation speed are 0–300 N m and 0–6000 rpm respectively, with accuracies of ±0.5%. The volumetric flow rate of the working fluid is regulated by adjusting the feed pump capacity in the range of 0–100%. During the experiment, the temperature and pressure data are automatically collected by Agilent 34980A data acquisition instrument. The power, rotation speed and torque are monitored by CYB-803S intelligent torque meter. All the output signals of the experimental data are transported to a computer and restored as a function of time there. The ORC system, after being evacuated, is charged with pure R245fa. The initial amount of the charged R245fa is 32.6 kg. During the test, the heating temperature of the boiler is fixed at 90 °C to provide a steady low-temperature heat source. Both of the hot water pump and the cooling water pump are maintained at their full capacities. A variety of working conditions are obtained by adjusting the feed pump capacity and the generator load. 4. Thermodynamic analysis In addition to output power and rotation speed, the efficiency is also an important metrics in performance evaluation of an expander. A thermodynamic analysis based on state properties is carried out to calculate the expander efficiency and the cycle efficiency defined in Eqs. (1)–(3). State properties for R245fa are calculated using REFPROP 8.0 [16].

Expander isentropic efficiency :gs ¼

hEXP;I  hEXP;O 0 hEXP;I  hEXP;O

ð1Þ

Expander mechanical efficiency : gm ¼

Pe ðhEXP;I qEXP;I  hEXP;O qEXP;O ÞV a

Cycle efficiency :gc ¼

hEXP;I  hEXP;O hEVP;I  hEVP;O

ð2Þ

ð3Þ

0

where hEXP;O is the ideal enthalpy of R245fa at the expander outlet, kJ/kg. The expander displacement can be calculated according to the internal geometrical structure of the expander. As shown in Fig. 3, the sectional area of the working chamber formed by the cylinder, the piston and the sliding vane is calculated using Eqs. (4) and (5).

1 AðhÞ ¼ 2

Z

h

2

ðR2  l Þdu

Taking the thickness of the sliding vane into account, the volume of the working chamber is calculated using Eqs. (6) and (7).

1 VðhÞ ¼ AðhÞH  zBH 2

ð6Þ

 qffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi  r h¼R 1 1 cos h  ð1  r=RÞ2 cos2 h þ 2r=R  1 R

ð7Þ

The theoretical displacement of the expander is calculated as follows:

V t ¼ nðV h¼108  V h¼28 Þ Then the actual displacement is obtained as follows:

V a ¼ gV V t

5. Experimental results and discussion The experimental data shown in this section were gained in Tianjin University, China. The experiment were carried out from 16:22 to 18:25 PM, lasted for about 2 h. Since the expander could be damaged by working fluid with high-speed during startup, the system ran at small flow rates for a proper warm-up from 16:22 to 16:38 PM. Then, the R245fa flow rate was increased gradually by adjusting the feed pump capacity. The data shown in the following figures are the dynamic part of the test. Fig. 6 shows the expander rotation speed and the R245fa volume flow rate varying with time respectively. The R245fa volume flow rate was controlled by the diaphragm metering pump. The feed pump capacity was initially set as 20% during startup. Afterwards, the pump capacity was increased step by step by 10%, and the adjustments were made by hand at 16:44, 17:13 and 17:36 PM. The heating temperature of the boiler was set as 90 °C. As shown in Fig. 6, the expander rotation speed increases, on the whole, with the increasing of R245fa volume flow rate, though some fluctuations caused by the variation of the generator load can be observed. As the volume flow rate increases from 2.1 l/ h to 5.25 l/h, the maximum rotation speed increases from 560 rpm to 780 rpm. An abnormal phenomenon that the expander rotation speed would drop suddenly as soon as the feed pump capacity was enlarged was observed during the test. This phenomenon may be attributed partly to that the flow area diameter of the pipe is reduced suddenly at the expander inlet, leading to the increase of flow velocity and the decrease of pressure. This throttling effect will be enhanced at the moment when the feed pump capacity is enlarged; as a result, the expander inlet pressure drops and the rotation speed decreases correspondingly.

Table 3 Uncertainties of efficiencies.

ð4Þ

ð5Þ

ð9Þ

where gv is the volumetric efficiency of the expander. A 70% volumetric efficiency was measured during an early test using compressed air. During the experiment, the major error sources are measurement accuracy and error resulting from data logging and reading by the computer. A multiple-sample uncertainty analysis is conducted for the experimental results. As shown in Table 3, the uncertainties of calculated efficiencies are less than ±5%, which is acceptable for the experimental research.

0

qffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 2 l ¼ ðR  rÞ cosðh  uÞ þ r 2  ðR  rÞ2 sin ðh  uÞ

ð8Þ

Unit Accuracy Uncertainty (%)

T

PEXI

PEXO

Pe

gs

gm

gc

°C ±0.2

MPa 0.003

MPa 0.001

kW 0.001

%

%

%

0.93–1.1

1.1–1.6

3.5–4.4

1270

N. Zheng et al. / Applied Energy 112 (2013) 1265–1274

7.5

800 750

Rotation speed Volume flow rate

7.0 6.5 6.0

650

5.5

600

5.0 550 4.5 500

4.0

450

Flow rate (l/min)

Rotation speed (rpm)

700

3.5

400

3.0

350

2.5 2.0

300 16:45:02

16:53:09

17:07:35

17:18:18

17:30:05

17:36:52

Time (hh:mm:ss) Fig. 6. Expander rotation speed and the R245fa volume flow rate with time.

Fig. 7a illustrates the change in the expander inlet temperature and pressure over time. The expander inlet temperatures changed between 76.5 °C and 87.7 °C, though both the heating temperature of the boiler and the hot water pump capacity remained unchanged during the test. The temperature fluctuations were attributed to the variation of the R245fa flow rate and the evaporating pressure. The expander inlet pressure which was also affected by the flow rate fluctuated violently between 0.622 MPa and 0.835 MPa, as shown in Fig. 7a. The change in expander outlet temperature and pressure over time is shown in Fig. 7b. As shown in Fig. 7b, the expander outlet temperatures changed between 42.1 °C and 64.7 °C, and the corresponding pressures changed between 0.115 MPa and 0.185 MPa. Overall, the expander outlet temperature increased as the inlet temperature increased and decreased as the inlet temperature droped. Unlike the expander outlet temperature, the expander outlet pressure held quite steady, regardless of the inlet pressure fluctuation. The stability of the expander outlet pressure could be attributed to the enough cooling capacity of the condenser. During the test, the condensation temperature was kept at around 23.0 °C. Fig. 8 shows the variation of shaft power versus rotation speed under various working conditions. As previously mentioned, the feed pump capacity was increased step by step by 10%. After the pump capacity was increased, a variety of expander rotation speed was obtained by adjusting the generator load. As shown in Fig. 8, the expander has a speed range of 320–780 rpm and an output power range of 50–350 W. It is worth noting that the maximum output power (350 W) did not correspond to the maximum rotation speed (780 rpm), instead, the maximum output power was obtained at lower rpm (660–670 rpm). This phenomenon indicates that the energy conversion efficiency decreases as the rotation speed increases, which will be discussed in detail in the following part. Fig. 8 also shows that a group of regression lines with different slopes can be plotted through the experimental data points. The slopes of the regression lines stand for the different output torques of the expander, which can be changed by adjusting the generator load. When the generator load is kept unchanged, the expander output power will increase with the increase of rotation speed in a certain range. The expander stability is affected by many internal and external factors, and the inlet pressure is decisive. Fig. 9 shows the curves of expander inlet pressure and rotation speed versus time from 16:52

to 17:12 PM respectively. The feed pump capacity was set as 30% and the expander rotation speed was controlled by adjusting the generator load. As shown in Fig. 9, the expander inlet pressure has a variation tendency similar to that of rotation speed. This indicates that higher rotation speed will be achieved by further increasing the inlet pressure. When the torque is fixed, the increase of rotation speed certainly leads to the growth of power output. In practice, however, the flow resistance loss and the internal leakage loss of the expander will also increase markedly with the increase of rotation speed. Hence, theoretically there exists a maximum power output as the inlet pressure increases gradually. Limited by the temperature of heat source, higher expander inlet pressures were not achieved during the test. Fig. 10 shows the variation tendency of output power and rotation speed versus the pressure difference through the expander respectively. The feed pump capacity was set as 50%. As shown in Fig. 10, the variation tendency of output power is similar to that of the rotation speed before the pressure difference reaches about 0.613 MPa. Both the output power and the rotation speed increase markedly, while their growth rates gradually slow down as the pressure difference increases. Afterwards, the rotation speed seems to experience another round of growth, whereas the output power peaks at about 0.35 kW and remains relatively stable as the pressure difference further increases to 0.642 MPa. This phenomenon indicates that the energy conversion efficiency decreases as the pressure difference increases. As mentioned in Section 2, appropriate clearances are left between the kinematic pairs such as sliding vane-piston, sliding vane–sliding slot, and piston–cylinder. The leakage amount of working fluid through these clearances will be increased by enlarging the pressure difference. The leakage will lead to the loss of flow, the decrease of inlet pressure and the increase of the exhaust temperature, and will finally reduce the expander efficiency. In addition to leakage, the flow resistance which reduces the fluid pressure and density also contributes to the decrease of expander efficiency. Since the flow resistance is proportional to the square of the velocity, the energy loss caused by flow resistance will increase markedly as the expander rotation speed rises. Fig. 11 shows the variation of efficiency versus rotation speed. As shown in Fig. 11, the isentropic efficiency remains relatively steady at around 40% over a speed range of 400–780 rpm, though a peak value of 43.3% is achieved at 680 rpm. When applying an

1271

N. Zheng et al. / Applied Energy 112 (2013) 1265–1274

1.10

100 95

1.00

Expander inlet pressure

0.95 85

0.90 0.85

80

0.80

75

0.75

Pressure (MPa)

Temperature ( )

90

1.05

Expander inlet temperature

70 0.70 65

0.65

60

0.60 17:00:19

17:13:58

17:30:59

17:45:29

17:59:28

18:14:12

Time (hh:mm:ss)

(a)

0.7

80

70

Expander outlet temperature Expander outlet pressure

0.6

0.4 50 0.3 40 0.2 30

Pressure (MPa)

Temperature ( )

0.5 60

0.1

0.0

20 17:02:19

17:15:37

17:33:58

17:49:49

18:04:50

Time (hh:mm:ss)

(b)

Fig. 7. (a) Change in expander inlet temperature and pressure over time. (b) Change in expander outlet temperature and pressure over time.

expander in a micro-CHP system, the expander rotation speed has to change frequently to meet the user demand. Hence, it is important for expanders to perform steadily at different power outputs. Unlike the isentropic efficiency, the mechanical efficiency is affected markedly by the rotation speed. Specifically, the mechanical efficiency first increases and then decreases with the increase of rotation speed, and reaches the peak value (about 44%) at 645 rpm. The relatively dramatic changes in mechanical efficiency may be attributed to the fluctuation of the volume flow rate of the working fluid. In this study, the Rankine cycle efficiency is simplified as the ratio between the enthalpy drop through expander and the change in enthalpy across the evaporator. As a result, the influence of the fluctuant R245fa flow rate on the cycle efficiency can be diminished. As shown in Fig. 11, the Rankine cycle efficiency is between 5% and 6% within the whole rotation speed range. In this study, the power consumption of the R245fa feed pump is not considered in the Rankine efficiency, because the feed pump applied in

the prototype can be powered by solar energy and wind energy [17,18]. The expander isentropic efficiency obtained in the experiment (43%) is a bit lower than that of scroll expanders (48–65%), and the overall efficiency is also low to be commercially viable. Limitation of the efficiency is partly explained by the low temperature of the heat source, which can be rectified by using higher temperature heat sources. Compared with [2], it can be concluded that a five percent increase in expander efficiency will be achieved by increasing the heat source temperature by about 20 °C. Beyond that, analyses show that the improvements in the rolling-piston expander design could improve the expander’s isentropic efficiency substantially [19]. During the test, the exhausted vapor at the expander outlet still possesses a high-temperature level (42.1–64.7 °C), thus there is much scope to improve the cycle efficiency by introducing internal heat exchanger into the Rankine cycle system [20,33].

1272

N. Zheng et al. / Applied Energy 112 (2013) 1265–1274

0.35

0.30

Experimental data point

Shaft power (kW)

0.25

0.20

0.15

0.10

0.05

0.00 300

350

400

450

500

550

600

650

700

750

800

Rotation Speed (rpm) Fig. 8. Change of shaft power with rotation speed.

700

0.82 Pressure Rotation speed

0.78

690 680

0.76

670

0.74

660

0.72 650 0.70 640

0.68

630

0.66 0.64

620

0.62

610

0.60 16:50

Expander rotation speed (rpm)

Expander inlet pressure (MPa)

0.80

600 16:53

16:56

16:59

17:02

17:05

17:08

17:11

Time (hh:mm) Fig. 9. Change in expander inlet pressure and rotation speed versus time.

Certainly, the efficiency will be affected by the vapor temperature and pressure greatly, and the efficiency could be improved by increasing the maximum temperature and pressure in the cycle. It should be noted that higher inlet pressure will also increase the flow resistance loss and the internal leakage loss of the expander, which reduces the expander efficiency and hence the cycle efficiency. Theoretically, there exists an optimal pressure and temperature at the expander’s inlet that leads to the maximum expander efficiency and maximum overall efficiency. However, higher pressure and temperature at the expander’s inlet were not achieved during the test because of the limited temperature of heat source. Thus the optimal pressure and temperature were not achieved experimentally.

der under varying conditions is verified. The experimental results indicate the following:

6. Conclusions

It should be noted that the R245fa rolling-piston expander technology is far from mature and still in a relatively early stage. Since no detailed R245fa rolling-piston expander model has been developed by far, the potential of improvement for

In this study, dynamic test of a kilowatt-sized rolling-piston expander is conducted and the performance of the proposed expan-

1. The proposed R245fa rolling-piston expander works stably during the test, with a maximum shaft power output of 0.35 kW and a steady isentropic efficiency of 40%. The Rankine cycle efficiency is steady between 5% and 6%. 2. The expander mechanical efficiency varies with expander rotation speed. A maximum mechanical efficiency of 44% is achieved at 645 rpm. 3. The expander rotation speed is affected by the flow rate of working fluid. As the flow rate increases from 2.1 l/min to 5.25 l/min, the maximum rotation speed increases from 560 rpm to 780 rpm.

1273

N. Zheng et al. / Applied Energy 112 (2013) 1265–1274

680

0.35 Power Rotation speed

640

600

560 0.25 520

480

0.20

Rotation Speed (rpm)

Power (kW)

0.30

440

0.15 0.40

400 0.45

0.50

0.60

0.55

0.65

Pressure Drop in Expander (MPa) Fig. 10. Variation tendency of output power and rotation speed versus the pressure difference.

60

ηR ηs

50

ηm

Efficiency (%)

40

30

20

10

0 400

450

500

550

600

650

700

750

800

Rotational Speed (rpm) Fig. 11. Variation of efficiency with rotation speed.

the expander’s efficiency cannot be determined. But in view of that the maximum theoretical efficiency of a CO2 rolling-piston expander is 60%, we think there is plenty of scope to improve the efficiency of the R245fa rolling-piston expander. Further work will be conducted with a view to increasing the expander output power by increasing the expander efficiency, reducing heat and pressure losses in the flow path, and optimizing the system layout and components. Acknowledgment The authors would like to acknowledge the financial support provided by the Program of National Nature Science Foundation of China (51276123).

References [1] Schuster A, Karellas S, Kakaras E, Spliethoff H. Energetic and economic investigation of Organic Rankine Cycle applications. Appl Therm Eng 2009;29:1809–17. [2] Wang XD, Zhao L, Wang JL, Zhang WZ, Zhao XZ, Wu W. Performance evaluation of a low-temperature solar Rankine cycle system utilizing R245fa. Sol Energy 2010;84:353–64. [3] Zhang SJ, Wang HX, Guo T. Performance comparison and parametric optimization of subcritical Organic Rankine Cycle (ORC) and transcritical power cycle system for low-temperature geothermal power generation. Appl Energy 2011;88:2740–54. [4] Qiu GQ, Shao YJ, Li JX, Liu H, Riffat SB. Experimental investigation of a biomassfired ORC-based micro-CHP for domestic applications. Fuel 2012;96: 374–82. [5] Wang DX, Ling X, Peng H. Performance analysis of double Organic Rankine Cycle for discontinuous low temperature waste heat recovery. Appl Therm Eng 2012;48:63–71.

1274

N. Zheng et al. / Applied Energy 112 (2013) 1265–1274

[6] Boretti A. Recovery of exhaust and coolant heat with R245fa Organic Rankine Cycles in a hybrid passenger car with a naturally aspirated gasoline engine. Appl Therm Eng 2012;36:73–7. [7] Lemort V, Quoilin S, Cuevas C, Lebrun J. Testing and modeling a scroll expander integrated into an Organic Rankine Cycle. Appl Therm Eng 2009;29: 3094–102. [8] Quoilin S, Lemort V, Lebrun J. Experimental study and modeling of an Organic Rankine Cycle using scroll expander. Appl Energy 2010;87:1260–8. [9] Liu H, Qiu G, Daminabo F, Shao YJ, Riffat SB. Preliminary experimental investigations of a biomass-fired micro-scale CHP with Organic Rankine Cycle. Int J Low Carbon Technol 2010;5:81–7. [10] Qiu GQ, Liu H, Riffat S. Expanders for micro-CHP systems with Organic Rankine Cycle. Appl Therm Eng 2011;31:3301–7. [11] Pei G, Li J, Li Y, Wang D, Ji J. Construction and dynamic test of a small-scale Organic Rankine Cycle. Energy 2011;36:3215–23. [12] Wang H, Peterson R, Harada K, Miller E, Ingram-Goble R, Fisher L, et al. Performance of a combined Organic Rankine Cycle and vapor compression cycle for heat activated cooling. Energy 2011;36:447–58. [13] Aoun. Micro combined heat and power operating on renewable energy for residential building: Ecole des Mines de Paris; 2008. [14] Zha ST. Study and development of CO2 transcritical expander. Ph.D. Dissertation. China: Tianjin University; 2002. [15] Yang JL, Ma YT, Li MX, Guan HQ. Performance investigation of transcritical carbon dioxide two-stage compression cycle with expander. Energy 2005;32:237–45. [16] Lemmon EW, McLinden MO, Huber ML. NIST reference fluid thermodynamic and transport properties – REFPROP. 8.0 ed: NIST Standard Reference, Database 23; 2002. [17] Nguyen VM, Doherty PS, Riffat SB. Development of a prototype lowtemperature Rankine cycle electricity generation system. Appl Therm Eng 2001;21:169–81. [18] Delgado-Torres AM. Solar thermal heat engines for water pumping: an update. Renew Sustain Energy Rev 2009;13:462–72. [19] Guan HQ, Ma YT, Li MX. Some design features of CO2 swing piston expander. Appl Therm Eng 2006;26:237–43.

[20] Yari M. Performance analysis of the different Organic Rankine Cycles (ORCs) using dry fluids. Int J Energy 2009;6:323–42. [21] Saitoh T, Yamada N, Wakashima S. Solar Rankine cycle system using scroll expander. J Environ Eng 2007;2:708–19. [22] Peterson RB, Wang H, Herron T. Performance of small-scale regenerative Rankine power cycle employing a scroll expander. Proc Inst Mech Eng, Part A: J Power Energy 2008;222:271–82. [23] Kane M. Small hybrid solar power system. Energy 2003;28:1427–43. [24] Yamamoto T, Furuhata T, Arai N, Mori K. Design and testing of the Organic Rankine Cycle. Energy 2001;26:239–51. [25] Liu H, Shao YJ, Li JX. A biomass-fired micro-scale CHP system with Organic Rankine Cycle (ORC)e thermodynamic modelling studies. Biomass Bioenergy 2011;35:3985–94. [26] Clemente S, Micheli D, Reini M, Taccani R. Energy efficiency analysis of Organic Rankine Cycles with scroll expanders for cogenerative applications. Appl. Energy 2012;97:792–801. [27] Fiaschi D, Manfrida G, Maraschiello F. Thermo-fluid dynamics preliminary design of turbo-expanders for ORC cycles. Appl. Energy 2012;97:601–8. [28] Wang H, Peterson RB, Herron T. Experimental performance of a compliant scroll expander for an Organic Rankine Cycle. Proc Inst Mech Eng, Part A: J Power Energy 2009;23:863–72. [29] Guzovic Z, Majcen B, Cvetkovic S. Possibilities of electricity generation in the Republic of Croatia from medium-temperature geothermal sources. Appl Energy 2012;98:404–14. [30] Quoilin S, Aumann R, Grill A, Schuster A, Lemort V, Lemort V. Hartmut Spliethoff, dynamic modeling and optimal control strategy of waste heat recovery Organic Rankine Cycles. Appl Energy 2011;88:2183–90. [31] Zhang HG, Wang EH, Fan BY. A performance analysis of a novel system of a dual loop bottoming Organic Rankine Cycle (ORC) with a light-duty diesel engine. Appl Energy 2013;102:1504–13. [32] Mago PJ, Luck R. Evaluation of the potential use of a combined micro-turbine Organic Rankine Cycle for different geographic locations. Appl Energy 2013;102:1324–33. [33] Liu B, Rivière P, Coquelet C, Gicquel R, David F. Investigation of a two stage Rankine cycle for electric power plants. Appl Energy 2012;100:285–94.