Energy 91 (2015) 324e333
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Energy journal homepage: www.elsevier.com/locate/energy
Experimental investigation on a small pumpless ORC (organic rankine cycle) system driven by the low temperature heat source P. Gao, L.W. Wang*, R.Z. Wang, L. Jiang, Z.S. Zhou Institute of Refrigeration and Cryogenic, Key Laboratory of Power Machinery and Engineering of MOE, Shanghai Jiao Tong University, Shanghai 200240, China
a r t i c l e i n f o
a b s t r a c t
Article history: Received 6 March 2015 Received in revised form 5 July 2015 Accepted 15 August 2015 Available online xxx
A small pumpless ORC (organic rankine cycle) system with different scroll expanders modiﬁed from compressors of the automobile air-conditioner is established, and the refrigerant R245fa is chosen as the working ﬂuid. Different hot water temperatures of 80, 85, 90 and 95 C are employed to drive the pumpless ORC system. Experimental results show that a maximum shaft power of 361.0 W is obtained under the hot water temperature of 95 C, whereas the average shaft power is 155.8 W. The maximum energy efﬁciency of 2.3% and the maximum exergy efﬁciency of 12.8% are obtained at the hot water temperature of 90 C. Meanwhile a test rig for investigating the mechanical loss of the scroll expander is established. The torque caused by the internal mechanical friction of the expander is about 0.4 N m. Additionally, another scroll expander with a displacement of 86 ml/r is also employed to investigate how scroll expander displacement inﬂuences the performance of the pumpless ORC system. Finally, the performance of the pumpless ORC system is compared with that of the conventional ORC system, and experimental results show that the small pumpless ORC system has more advantages for the low-grade heat recovery. © 2015 Elsevier Ltd. All rights reserved.
Keywords: Pumpless ORC (organic rankine cycle) Scroll expander Mechanical loss Energy efﬁciency
1. Introduction As one highly desirable approach to recover the low-grade heat, ORC (organic rankine cycle) has the characteristic of the simplicity, ﬂexibility and relatively low driving temperature . But the relatively low net power efﬁciency of ORC system always plagues worldwide researchers. The basic method to improve the power efﬁciency is to optimize the performance of each component of the ORC system. For the evaporator, the PPTD (pinch point temperature difference) was employed by Li et al.  and Guo et al.  to determine the optimal evaporating pressure and evaporating temperature. For the condenser, Qiu et al.  developed a biomass-ﬁred ORC-based micro-CHP system, and used the cooling water leaving the condenser for the domestic washing and under-ﬂoor heating. Both evaporators and condensers can be easily gotten and chosen from the HVAC industries at a reasonable price, but for the expander and working ﬂuid pump, they are not easily gotten and chosen. The expander is the core part of the ORC system. Different kinds of expanders such as scroll expanders , the radial turbines ,
* Corresponding author. Tel.: þ86 21 34208038; fax: þ86 21 34206814. E-mail address: [email protected]
(L.W. Wang). http://dx.doi.org/10.1016/j.energy.2015.08.076 0360-5442/© 2015 Elsevier Ltd. All rights reserved.
and the single screw expanders  are utilized to convert the thermal energy to the mechanical shaft power. Especially, scroll expanders have been widely adopted in small ORC systems. Generally, scroll expanders are modiﬁed from scroll compressors for the vapor compression cycles and the compressed air production. Carlos et al.  proposed a semi-empirical model of the scroll expander modiﬁed from the compressor used in the automotive refrigeration system, and this model could predict the mechanical power, the exhaust temperature and the supply mass ﬂow rate with good accuracy. Declaye et al.  made an open-drive and oil-free scroll expander modiﬁed from the air scroll compressor and had successfully solved the absence of tightness which is the major drawback of an open-drive air scroll expander. According to the scroll expander model developed by Lemort  employing HCFC123, Giuffrida  presented a performance simulation tool of a scroll expander for any working ﬂuid. But how scroll expander displacement inﬂuences the performance of the ORC system has not been investigated before. For the working ﬂuid pump, the refrigerant is pressurized here from the condensing pressure to the evaporating pressure, so a certain amount of electricity is naturally consumed, which cannot be neglected especially for a small ORC system. Consequently the net power efﬁciency is greatly reduced, and meanwhile the
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Nomenclature Q t M C T m h W n N
h T0 E
heat transfer rate, W time, s total mass, kg speciﬁc heat, kJ/(kg K) temperature, K mass ﬂow rate, kg/s speciﬁc enthalpy, J/kg power, W rotational speed, r/min torque, N m efﬁciency environment temperature, C exergy, W
working ﬂuid pump also seriously affects the compactness and the cost of the ORC system . For solving this problem Li et al.  developed a novel gravity driven ORC for small-scale cogeneration applications, but the lowest required height for pressurization is about 20.9 m with the working ﬂuid PF5060. Although the working ﬂuid pump is replaced, the enough room is necessary for the novel gravity driven ORC. Yamada et al.  developed a new PRC (pumpless rankine-type cycle), which mainly consisted of a scroll expander, two heat exchangers, and switching valves for the expander and heat exchangers. In their initial experiments an expansion nozzle was employed to simulate the expander for clarifying the feasibility of the PRC. Experimental results conﬁrmed that the proposed cycle had the potential to produce power. Then an experimental system of PRC with the scroll expander was established, and corresponding experiments were also conducted. The PRC system had the characteristics of the compactness, low cost and relatively high net power efﬁciency. But the maximum shaft work obtained was very small, which was only 50 W. They also didn't investigate the performance of the system under different heat source temperatures, as well as didn't test the performance of the system with the scroll expanders of different displacements. In order to investigate the performance of the pumpless ORC thoroughly, in this paper a pumpless ORC system with the scroll expander modiﬁed from the automobile air-conditioner compressors is established, and the expanders with the displacement of 66 ml/r and 86 ml/r are tested. This pumpless ORC system established in this paper is much bigger than that was established by Yamada et al. , so more reliable data is expected to be gotten. Two tank-type heat exchangers are utilized in this system as the evaporator or condenser. Although GWP (global warming potential) of the refrigerant R245fa is 820, its ODP (ozone depletion potential) is 0 and it has reasonable performance for the lowtemperature heat source , thus it is employed as the working ﬂuid. The performance of the pumpless ORC system is tested under different hot water temperatures ranging from 80 to 95 C, and the mechanical loss of the scroll expander is also investigated based on a corresponding test rig. 2. System description The pumpless ORC system is shown in Fig. 1a, which consists of two same heat exchangers, a scroll expander, four refrigerant valves, eight water valves, an electricity generator and other auxiliary equipments. A combination of four valves serves as a two-
friction force, N equivalent lever arm, m
Subscripts h heat ref refrigerant exc heat exchanger m metal w water hw hot water eva evaporator ave average exp expander mech mechanical s isentropic
position four-way valve. Compared with the conventional ORC system, in the pumpless ORC system the working ﬂuid pump is replaced by three four-way valves. Additionally hot water/cooling water is used as the heat transfer ﬂuid for evaporation/condensation processes. The working process of the pumpless ORC system consists of the preheating process and the power generating process, and the system outputs the power only at the second process. The detailed working principles are described as the follows (Fig. 1a and Fig. 1b): (a) HX1 works as an evaporator, and HX2 works as a condenser. This process is deﬁned as phase 1. Water valves V2, V4, V5 and V7 are open, and all other valves including refrigerant valves RV1, RV2, RV3 and RV4 are closed. The hot water from the heat source ﬂows into the HX1 to preheat HX1, which is an isochoric heating process. With time elapsing, the pressure of the HX1 increases gradually until it reaches an almost constant value that is quite close to R245fa saturation pressure corresponding to the hot water temperature. Simultaneously the cooling water from the cooling tower ﬂows into the HX2 to cool HX2. (b) When the pressure of the HX1 increases to an almost constant value, RV1 and RV4 are opened. Then the high-pressure and high-temperature R245fa from the HX1 ﬂows into the expander and expands, which is 1e2 in Fig. 1b. The expander outputs shaft power until there is no pressure difference between the HX1 and HX2. At this moment the heating process inside HX1 is an isobaric heating process, which is 4e1 in Fig. 1b. The exhaust enters the HX2 and is condensed into the saturated liquid at state 3 in Fig. 1b, and the condensation process is also isobaric process of 2e3 in Fig. 1b. When the expander doesn't output shaft power, RV1, RV4, V2, V4, V5 and V7 are closed. This process is the power generating process. (c) HX1 works as a condenser, and HX2 works as an evaporator. This process is deﬁned as phase 2. Water valves V1, V3, V6 and V8 are opened, so hot water begins to heat HX2 and cooling water begins to cool HX1. The following processes are similar to those described in the (a) and (b). To measure the shaft power of the expander, a torque meter of TQ-669 is installed between the expander and the electricity generator, and the measuring precisions of the rotational speed and torque are ±(0.04%þ2) and 0.3%R. Temperature sensors of PT100 and pressure transmitters of YSZK-31 are mounted at the
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Fig. 1. Schematic diagram of (a) the pumpless ORC system; (b) T-s diagram of pumpless ORC.
corresponding positions as shown in Fig. 1a, and their accuracies are ±0.1 C and 0.3%R respectively. The ﬂow rate of hot water is measured by the turbine ﬂowmeter of LWGY-B15L, and its accuracy is 0.5%.
hexp ¼ hexp;mech
3. Performance calculation
In the conventional ORC system, if the heat loss transferred from heat exchanger to surroundings is neglected, under the stable phases heat source supplying for ORC system only includes the heat transferred to R245fa for evaporation because the heat exchanger is maintained at a constant temperature. Different from the conventional ORC system, the heat input Qh for the pumpless ORC system is composed of two parts, i.e. the heat Qref transferred to R245fa for evaporation and the heat to warm up the heat exchanger. It is
Qh ¼ Qref þ Z
1 Mexc Cm $DTm tcycle
Q ref ¼
t Zpower 7 1 6 4Mref heva;mid hliq;sat þ mref heva;out heva;mid dt5
(2) where heva,mid is the enthalpy of the ﬂuid at the point 4 in Fig. 1b; heva,out is the enthalpy of the ﬂuid at the point 1 in Fig. 1b; hliq,sat is the enthalpy of the saturated liquid at the point 3 in Fig. 1b; tpower is the time of the power generating process. The average outputting shaft power of the expander is
1 tcycle 1 tcycle 1 tcycle
Wdt 0 tcycle Z
hexp;mech mref heva;out hexp;out dt
0 tcycle Z
2p n N dt 60
The scroll expander efﬁciency is
The exergy efﬁciency of the pumpless ORC system is
Wave ¼ E
where is the thermal exergy, Thw,ave is average temperature of the hot water. The ideal energy efﬁciency of the pumpless ORC system is deﬁned as tcycle;ORC
hinternal ¼ Z0 t cycle;ORC 0
where tcycle is the cycle time. The heat Qref is
The energy efﬁciency of the pumpless ORC system is
Cw $mw $ Thw;in Thw;out dt
heva;out hexp;out heva;out hexp;out;s
where heva,out represents the outlet enthalpy of the evaporator, and hliq,sat represents the initial enthalpy of the working ﬂuid inside the evaporator, which has not been heated yet. And heva,out e hexp,out represents the enthalpy difference between the expander inlet and the expander outlet. The ideal energy efﬁciency of the pumpless ORC system refers to the maximum energy efﬁciency that the pumpless ORC system can obtain, thus the mechanical friction loss of the expander and metallic thermal capacity of heat exchangers are not considered in the equation.
4. Experimental results and discussion The photo for the established pumpless ORC system is shown in Fig. 2. The water boiler heated by the electric heater is utilized to simulate the low-grade heat source. The cooling water from a cooling tower takes the condensing heat of the refrigerant R245fa away. The scroll expander modiﬁed from the compressor of automobile air-conditioner is chosen for the power generating process, and its displacement is 66 ml/r in compressor mode. Two same tank-type heat exchangers from HVAC industry are directly used as the evaporator and condenser, respectively. This tank-type heat exchanger could storage about 5.1 kg refrigerant liquid. The length of the copper spiral tube inside the tank-type heat exchanger is 5.1 m, and its size is F19 1.5 mm. Four refrigerant valves and eight
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Fig. 4. Local resistance loss of T-junctions.
Fig. 2. The pumpless ORC system.
water valves are employed to change the ﬂow direction of the refrigerant or water. In order to determine the optimal operating condition of the pumpless ORC system, different hot water temperature ranging from 80 to 95 C is studied and different scroll expanders are employed. The inﬂuence of the metallic thermal capacity of HXs and the internal mechanical friction loss of the scroll expander on the pumpless ORC system performance is investigated in detail. Meanwhile the performance of the pumpless ORC system is compared with that of the conventional ORC system. 4.1. Performance inﬂuenced by different hot water temperatures Different hot water temperatures of 80, 85, 90 and 95 C are employed to drive the pumpless ORC system to investigate how the performance of the pumpless ORC system is inﬂuenced by the hot water temperature, and the ﬂow rate of the hot water is 2.3 m3/h. The trends of the evaporating temperature and the expander outlet pressure are shown in Fig. 3 when the hot water temperature and the cooling water temperature are 80 C and 25 C. Experimental results show that the variation of evaporating pressure over time is almost same for the evaporators of HX1 and HX2. When the
preheating process begins, the evaporating pressure increases sharply to an almost constant value of 0.78 MPa at about 27 s. Then expansion process begins, the evaporating pressure decreases to a constant value of 0.65 MPa, which is maintained at about 36 s. After that it decreases again to a constant value of 0.47 MPa, which lasts for about 54 s until the expander stops working. There are two almost constant values, which is different from previous theoretical analysis shown in Fig. 1b. The difference is mainly due to the actual heat exchange between the refrigerant R245fa and the hot water within the HX. When the preheating process is completed, the evaporating pressure reaches 0.78 MPa corresponding to the saturated temperature of 79.53 C. At this moment the heat transfer temperature difference is less than 0.5 C. When the power generating process begins, the refrigerant R245fa needs adsorb a large amount of heat for evaporation. However the heat transfer area is constant, so lager heat transfer temperature difference is essential. As a result, the evaporating pressure decreases from 0.78 MPa to 0.65 MPa corresponding to the saturated temperature of 72.42 C. With time elapsing, the mass of the refrigerant R245fa within the HX decreases gradually, so the effective heat transfer area is reduced greatly. In order to meet the heat transfer requirement larger heat transfer temperature difference is necessary. Consequently, the evaporating pressure decreases again from 0.65 MPa to 0.47 MPa corresponding to the saturated temperature of 60.56 C. Fig. 3 also shows for the expander outlet pressure there is some distinction between two phases. This distinction is due to the asymmetric conﬁguration of the refrigerant piping between heat exchangers and the scroll expander. This local resistance loss of phase 1 is greater than that of phase 2, so the expander outlet pressure of phase 1 is higher than that of phase 2. The reason is as
Fig. 3. Hot water temperature, cooling water temperature, evaporating pressure and expander outlet pressure vs. cycle time.
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Fig. 5. Evaporating pressure and shaft power of the scroll expander of (a) 80 C hot water temperature; (b) 85 C hot water temperature; (c) 90 C hot water temperature; (d) 95 C hot water temperature.
follows: The local resistance of phase 1 contains four T-junctions shown in Fig. 4a, and the local resistance of phase 2 contains four Tjunctions shown in Fig. 4b. The coefﬁcient of local resistance of Fig. 4a is 1.3, and the coefﬁcient of local resistance of Fig. 4b is 0.1. Consequently the local resistance loss of phase 1 is greater than that of phase 2. The shaft power of the scroll expander under the hot water temperatures of 80, 85, 90 and 95 C is shown in Fig. 5, and the
corresponding evaporating pressures under different hot water temperatures are also shown in Fig. 5. After the power generating process begins, the shaft power increases quickly and then decrease gradually for both evaporators of HX1 and HX2. Here, when the HX1 works as the evaporator the working process is phase 1. Whereas when the HX2 works as the evaporator the working process is deﬁned as the phase 2. It can be found that the maximum shaft power of the phase 1 is lower than that of the
Fig. 6. Inlet temperature and outlet temperature of the hot water vs. time: (a) 80 C hot water temperature; (b) 85 C hot water temperature; (c) 90 C hot water temperature; (d) 95 C hot water temperature.
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Fig. 7. Hot water outlet temperature and shaft speed under the hot water inlet temperature of 80 C vs. time.
phase 2, which is due to that the lower pressure at the expander outlet of phase 2. I.e. the expander pressure ratio of the phase 1 is smaller than that of the phase 2, and consequently the enthalpy difference of phase 1 is lower than that of the phase 2. It should be noted that even if the evaporating pressure is constant, the shaft power still decreases with time, because the mass ﬂow rate of the refrigerant R245fa decreases gradually. Under the hot water temperature of 95 C, a maximum shaft power of 361.0 W is obtained, whereas an average shaft power is 155.8 W. From Fig. 5, it is also found that the cycle time decreases little by little with the increase of the hot water temperature. It is because higher temperature accelerates the heat transfer process, and the fast heat transfer process to the refrigerant R245fa causes a lager mass ﬂow rate, which is reﬂected in the higher shaft speed of the scroll expander. When the hot water temperature increases from 80 C to 95 C, average shaft speed increases gradually from 1559 rpm to 1603 rpm.
taken as the example, and reasons are as follows: First, at the beginning of the heating process the original cooling water inside HX2 needs to be discharged. Second the heat is consumed by the metallic thermal capacity of heat exchanger which needs to be heated from low temperature to high temperature. Additionally a certain amount of heat is consumed for heating up the sub-cooled refrigerant liquid of R245fa to saturated state. Fig. 6aed also show that when the power generating process occurs, the hot water outlet temperature drops slightly again. It is because a large amount of refrigerant R245fa that is 5.1 kg has the large heat transfer area and starts to evaporate. The large evaporation amount of refrigerant R245fa reﬂects in the sharp increase of the shaft speed. Hot water outlet temperature and shaft speed under the hot water temperature of 80 C are shown in Fig. 7. From the Fig. 7, it can be found that when the hot water outlet temperature drops, the shaft speed of the expander increases rapidly. With the working time elapses the amount of the refrigerant in the tanktype heat exchanger decreases, consequently the heat transfer area that the refrigerant contacts is less, thus the mass ﬂowrate of the evaporating refrigerant decreases, and the shaft power decreases accordingly. Based on Figs. 5 and 6, the average shaft power of the scroll expander and the average heat input of the hot water under different hot water temperatures are calculated, shown in Fig. 8a, both of them increase with hot water temperature, and the values are 105.1 We155.8 W and 5.5 kWe7.7 kW, respectively. Experimental results of energy and exergy efﬁciencies are shown in Fig. 8b. Both efﬁciencies increase with the increment of hot water temperature and then decrease slowly with the further increment. The maximum energy efﬁciency of 2.3% is obtained at a hot water temperature of 90 C, and the largest error of the energy efﬁciency is 5.6%. At same time the exergy efﬁciency is 12.8%. The error of the energy efﬁciency of the pumpless ORC system is calculated based on the equation as follows:
vﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃﬃ !2 !2 u 2 2 2 u vh vhenergy vhenergy vhenergy vhenergy energy Dn þ DN ¼t þ DThw;in þ DThw;out þ Dmw vn vN vThw;in vThw;out vmw
Fig. 6 shows the temperature variation of the hot water inlet and outlet. At the beginning of the heating process the hot water outlet temperature drops sharply. For example in Fig. 6a the temperature of the hot water decreases from 80 C to 45 C. In order to explain this phenomenon the heating process for HX2 is
The metallic thermal capacity of HXs and the inefﬁcient expander account for the relatively low energy efﬁciency of the pumpless ORC system. In order to estimate the inﬂuence of two parts, the ideal energy efﬁciency of the pumpless ORC system deﬁned by eq. (7) is calculated. Fig. 9 shows the enthalpy difference
Fig. 8. Performance of the pumpless ORC vs. hot water inlet temperature: (a) average heat input and average shaft power output; (b) energy efﬁciency and exergy efﬁciency.
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power of the scroll expander shown in Fig. 8b would increases from 105.1 W to 165.3 W. As a result, the energy efﬁciency would increases from 1.98% to 3.0% under the hot water temperature of 80 C.
Fig. 9. Enthalpy difference heva,out hliq,sat and heva,out hexp,out vs. cycle time.
heva,out hliq,sat and heva,out hexp,out under the hot water temperature of 80 C. Consequently, the ideal energy efﬁciency is 5.86%, which is higher than the experimental energy efﬁciency of 1.98%. The detailed analysis about the metallic thermal capacity of HXs and the inefﬁcient expander are as follows: (1) Scroll expander. There are four losses affecting the performance, and they are the pressure drop loss, the internal mechanical friction loss, the internal leakage loss and the heat transfer loss from the expander body surface, among which the mechanical friction loss plays an important role . According to the reference , for the open-drive scroll expanders the friction torque is not related with the rotational speed, and it should be constant. In order to investigate the mechanical friction loss, a corresponding test rig shown in Fig. 10a is established. An electromotor is employed to drive the scroll expander with the supply and exhaust ducts open, and a torque meter is installed between them to measure the torque and rotational speed. The internal mechanical friction torque is deﬁned as
N ¼F L
where F represents the friction force between two scroll wraps of the scroll expander, and L represents the equivalent lever arm. Experimental results indicate that the torque caused by the internal mechanical friction of the expander is about 0.4 N m. If the internal lubrication of the expander is under extremely good condition, the mechanical friction could be neglected and the shaft
(2) Metallic thermal capacity of HXs. The virtual energy efﬁciency of 3.0% is still lower than the ideal energy efﬁciency of 5.86%, which is mainly due to the big metallic thermal capacity of HXs. Before the power generating process, lots of heat from the hot water is used to warm up the heat exchanger. The total weight of one heat exchanger is 10.3 kg, and the weight of the copper spiral tube and the iron shell are 6.7 kg and 3.4 kg, respectively. Compared with the heat adsorbed by the refrigerant R245fa this part of the heat cannot be ignored because the refrigerant mass that the HX contains is relatively small, i.e. 5.1 kg. Based on those data, the ratio between the heat transferred to R245fa for evaporation and the heat to warm up the heat exchanger under the hot water temperature of 80 C is ﬁve to one. If a big system is employed and contains enough refrigerant R245fa, the ratio will be improved greatly, and consequently the heat consumed by the metallic thermal capacity of heat exchanger would decrease efﬁciently and subsequently the energy efﬁciency will be improved effectively.
4.2. Performance inﬂuenced by the expander displacement Under identical operating conditions and a hot water temperature of 85 C, another scroll expander with a displacement of 86 ml/r, as shown in Fig. 11a, is also employed to investigate how scroll expander displacement inﬂuences the performance of the pumpless ORC system. Fig. 11b shows that the system evaporating pressures for the systems with different expanders are almost same because the steam generating device for both systems is identical. Fig. 11c shows that the shaft power of the 86 ml/r expander is less than that of the 66 ml/r expander, which is mainly due to that the mechanical loss of the 86 ml/r expander is a little higher than that of the 66 ml/r expander. The torque caused by the internal mechanical friction of the 86 ml/r expander is also measured by the mechanical loss test rig shown in Fig. 10a, and is about 0.6 N m, which is higher than the loss of the 66 ml/r expander. It is because the equivalent lever arm L of the 86 ml/r expander is relatively larger than that of the 66 ml/r expander. At same time, it is found that the cycle time of the 86 ml/r expander is less than that of the 66 ml/r expander. The cycle time depends on the displacement and the rotational speed of the scroll expander. Generally speaking, if
Fig. 10. Mechanical loss test rig of the scroll expander: (a) experimental system; (b) experimental results.
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Fig. 11. Performance of the pumpless ORC system inﬂuenced by different scroll expanders: (a) photo of the scroll expanders; (b) evaporating pressure vs. cycle time; (c) shaft power vs. cycle time; (d) rotational speed vs. cycle time.
the amount of evaporated R245fa is the same, the larger displacement corresponds to the less cycle time and lower rotational speed shown in Fig. 11d. 4.3. Performance comparison of ORC system and pumpless ORC system The performance of the pumpless ORC system is compared with the performance of one conventional ORC system with the 66 ml/r expander in the lab to determine whether the small pumpless ORC system has advantages over the small conventional ORC system for the application in the low-grade heat recovery. The diagram of the conventional ORC system is shown in Fig. 12, which consists of an evaporator, a scroll expander, a condenser, a liquid storage tank, a metering pump, and an internal heat exchanger, etc. . For the conventional ORC system established in the lab, the performance of the system is tested under the different heat source temperature ranging from 85 C to 110 C. When the heat source temperature is below 95 C, the maximum energy efﬁciency is below 1.6%. So a higher heat source temperature of 105 C is chosen for the comparison. The corresponding experimental results are shown in Fig. 13. The maximum shaft power of 151.3 W is obtained at the evaporating pressure of 0.68 MPa. At this time, if the electricity consumed by the metering pump is not taken into account, the energy efﬁciency shown in Fig. 13a is 2.9%. The shaft power increases initially with the evaporating pressure and then decreases gradually with the further increase in the evaporating pressure. It is because the refrigerant volumetric ﬂowrate at the outlet of the metering pump decreases from 30.9 l/h to 15.4 l/h, and the sharp drop in volumetric ﬂowrate is caused by a rapid increase in the
pressure ratio of the metering pump. Both the volumetric ﬂow rate and the pressure ratio of the metering pump are shown in Fig. 13b. If the electricity consumed by the metering pump is taken into consideration, the net power output is almost negative, i.e. the power generation cannot balance the electricity consumption of the metering pump. Three two-position four-way valves in the pumpless ORC system play a similar role as the metering pump in the conventional ORC system. The electricity of about 100 W is consumed by three electric four-ways valves only occurring at the switch time (half cycle) that remains about for 6 s. The half cycle is 137 s, and the average power consumed by the valve is only 4.37 W, so it can be neglected. In addition, under the hot water temperature of 95 C shown in Fig. 5d, the shaft power of the expander higher than 300 W can last about 21 s in phase 2, the power generation during this time is enough to offset the electricity consumed by three four-ways valves. Consequently, the net energy efﬁciency of the pumpless ORC system is positive. 5. Conclusions In this paper, a pumpless ORC system with scroll expander modiﬁed from automobile air-conditioner compressor is established and investigated experimentally. The refrigerant R245fa is used as the working ﬂuid, and two types of expanders, i.e. expanders with 66 ml/r and 86 ml/r displacement are studied. Conclusions are as follows: (1) Different hot water temperatures ranging from 80 to 95 C are employed to drive the pumpless ORC system with 66 ml/r
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Fig. 12. The ORC of: (a) Schematic diagram; (b) experimental system.
Fig. 13. Performance of the ORC vs. evaporating pressure: (a) energy efﬁciency and shaft power; (b) volumetric ﬂowrate and pressure ratio of the pump.
P. Gao et al. / Energy 91 (2015) 324e333
scroll expander. Experimental results show that a maximum shaft power of 361.0 W is obtained under the hot water temperature of 95 C. The maximum energy efﬁciency of 2.3% and the maximum exergy efﬁciency of 12.8% are obtained at a hot water temperature of 90 C. The relatively low energy efﬁciency of the pumpless ORC system is mainly due to the metallic thermal capacity of heat exchanger and the internal mechanical friction loss of the scroll expander. The heat to warm up the heat exchange is about one ﬁfth of the heat transferred to R245fa for evaporation under the hot water temperature of 80 C. (2) The internal mechanical friction loss of the scroll expander is investigated in detail, and a corresponding test rig is established. The experimental result indicates that the torque caused by the internal mechanical friction between two scroll wraps is about 0.4 N m for the 66 ml/r expander. If the internal lubrication of the expander is under extremely good condition, the mechanical friction could be neglected and the energy efﬁciency would increase from 1.98% to 3.0% under the hot water temperature of 80 C. (3) Additionally, under a heat source temperature of 85 C, another scroll expander with a displacement of 86 ml/r is employed in the pumpless ORC system. The shaft power of the 86 ml/r expander is less than that of the 66 ml/r expander, which is mainly due to that the bigger mechanical friction torque of about 0.6 N m. At same time, it is found that the cycle time of the 86 ml/r expander is less than that of the 66 ml/r expander. (4) Experimental performance of the conventional ORC system and pumpless ORC system is compared. For the conventional small ORC system the net power generation is almost negative because the metering pump consumes large amount of electricity. As for the pumpless ORC system, the electricity consumed by switching valves could be neglected, so the net power generation is positive. As a result, compared with the conventional small ORC system, the small pumpless ORC system has advantage of higher efﬁciency for the low-grade heat recovery. Although the performance of the pumpless ORC is prospective, the unstable output of the mechanical energy, which is converted into the unstable electricity, will inﬂuence the usage of the energy somehow. The batteries that can storage the unstable electricity may be a solution for this problem, and it needs to be veriﬁed by more investigations in the future.
Acknowledgments This research was supported by the National Science Foundation of China for Excellent Young Scholars under contract number 51222601, the Program for New Century Excellent Talents in University by the Ministry of Education, China (NCET-11-0333), and the Shanghai Pujiang Program.
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