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Experimental study on low-temperature organic Rankine cycle utilizing scroll type expander
Applied Energy 155 (2015) 150–159
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Experimental study on low-temperature organic Rankine cycle utilizing scroll type expander Jen-Chieh Chang a, Tzu-Chen Hung b,⇑, Ya-Ling He c, Wenping Zhang d a
Graduate Institute of Mechanical and Electrical Engineering, National Taipei University of Technology, Taipei, Taiwan Department of Mechanical Engineering, National Taipei University of Technology, Taipei, Taiwan c Key Laboratory of Thermo-Fluid Science and Engineering of MOE, School of Energy and Power Engineering, Xi’an Jiaotong University, Xi’an, Shaanxi, PR China d College of Power and Energy Engineering, Harbin Engineering University, Harbin, Heilongjiang Province 150001, PR China b
h i g h l i g h t s Experimental studies of the ORC system operating with R245fa are presented. Scroll type expander was employed in the ORC system. The cycle was fueled with hot water (maximum 95.6 °C). Superheating improves isentropic efficiency of the expander.
a r t i c l e
i n f o
Article history: Received 6 February 2015 Received in revised form 11 May 2015 Accepted 25 May 2015
Keywords: Organic Rankine cycle Low-temperature Scroll expander Waste heat recovery Superheating
a b s t r a c t This paper focuses on experimental performance of an open-drive scroll type expander in an organic Rankine cycle (ORC) system. The expander was an originally oil-free scroll type air compressor with a built-in volume ratio of 4.05. The cycle used HFC-245fa as working fluid, and the loop has been mixed with a moderate concentration of refrigerant oil that circulated in the cycle. The experimental results of this study are divided into two main parts: first part focuses the experimental performance on the fixed superheating at the expander inlet with respect to various pressure differences of the system and rotational speeds of the expander. Second part involves various superheating at the expander inlet which was operated at fixed rotational speed and operating pressure difference of 5 bars and 6 bars. When the cycle was operated under fixed superheating conditions, the maximum cycle efficiency, expander efficiency and power output of the expander are 9.44%, 73.1% and 2.3 kW respectively. On the other hand, when the expander is operated in various superheating conditions, the expander and cycle efficiency simultaneously increase with the increasing of superheating. In addition, this paper not only focuses on the experimental results using the current expander, but also integrates the previous experimental data with present study to identify an appropriate scroll type expander with respect to various operating pressure differences for the heat source below 100 °C. Ó 2015 Elsevier Ltd. All rights reserved.
1. Introduction Organic Rankine cycle (ORC) is a promising machine for energy conversion from low-temperature heat sources. U.S. Department of Energy (U.S. DOE) has indicated that approximately 60% of low-temperature waste heat (below 232 °C) is not recovered, and this heat is in large quantities [1]. In order to convert low-temperature heat sources to useful work, Yamamoto et al. [2] experimentally demonstrated that the
⇑ Corresponding author. Tel.: +886 2 2771 2171x2021; fax: +886 2 2731 7191. E-mail address: [email protected] (T.-C. Hung). http://dx.doi.org/10.1016/j.apenergy.2015.05.118 0306-2619/Ó 2015 Elsevier Ltd. All rights reserved.
use of a low boiling point working fluid, such as R-123 utilized in the Rankine cycle, result in a better cycle performance than using water for the turbine inlet temperature below 120 °C. For this reason, the ORC systems have been widely tested for various applications. Qiu et al. [3] tested a biomass-fired CHP (combined heat and power) with an ORC for domestic application. Yagoub et al. [4] tested the ORC system driven by solar and natural gas in an office building, it reported that the system can effectively save the electricity and heat demand of the building. Wang et al. [5] tested both evacuated solar collectors and the plate collectors for absorbing solar energy in an ORC system, the result shows that the overall
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Nomenclature h _ m Q_
DP P s T _ W
enthalpy (kJ/kg) mass flow rate (kg/s) thermal power (kW) pressure difference of the system, P3 P1 (bar) pressure (bar) specific entropy (kJ/kg K) temperature (°C) work (kW)
Greek symbols g efficiency
thermodynamic power generation efficiency is 4.2%, when the evacuated solar collectors is utilized in the ORC, and with the condition of plat solar collector, it is about 3.2%. Wang et al. [6] tested a flat-plate collector for gathering solar energy in a regenerative ORC, and the result shows that both thermal and collector efficiencies could be improved significantly by adjusting the working fluid flow rate to an appropriate level. Zheng et al. [7] experimentally tested an ORC system for recovering the exhaust gas from a diesel engine, result shows that the maximum improvement of overall system efficiency was 1.53%. Endo et al. [8] incorporated an ORC in the automotive engine, they reported that the thermal efficiencies can be improved about 10%. ORCs also have been investigated to find a maximum work output from various combination thermodynamic cycles, such as recovering exhaust heat from the gas turbine [9], the high-temperature gas-cooled reactors [10], and the large marine diesel engine applications [11]. The applications of ORC system have a wide spectrum of recovering various waste heat sources or renewable energy. The key factor of the ORC is the selection of working fluids. Hung et al. [12,13] reported that the system efficiency increases and decreases for wet and dry fluids, respectively, and the isentropic fluids achieve an approximately constant value as increase the superheating of the expander inlet. Furthermore, the theoretical report has shown that wet fluids with very steep saturated vapor curves in T-s diagram have a better overall performance in energy conversion efficiencies than that of dry fluids [14]. ORCs are commonly operated at relatively low pressure differences in finite temperature ranges. Therefore, it is very important to reduce losses of the components of the system in order to achieve higher performance, especially in small-scale ORC systems. In the expansion machine, Bao et al. [15] summarized the advantages and disadvantages of various type expanders used in the ORC system. They indicated that the scroll and multi-vane expander are applicable to the power output smaller than 10 kW for the ORCs. Nevertheless, numerous publications mainly focused on ORCs with power output smaller than 5 kW in the laboratory tested. Table 1 summarizes recently experimental results which focus on the power output below 5 kW, without utilizing regenerator in the ORC system. It seems like that the positive-displacement expander such as scroll type expander is widely used for this scale of ORCs, because of the scroll machine can tolerate two-phase conditions [28], compact size, low cost, high efficiency and lower moving-parts. Bao et al. [15] reported that the capacity of the multi-vane expander could be reached to 10 kW. However, according to available literature, the maximum power output was only 1.72 kW [3]. Multi-vane expander has a simple structure, easier to manufacture and low cost, but shows a relatively low efficiency over positive-displacement expanders
Subscript cond ele exp evap H L in is out pp
condenser electricity expander evaporator hot, high low inlet isentropic outlet pump
(refer to Table 1), it might be a main reason why the experimental studies on these expansion machine are very few in the literature. Recently, it was claimed that the maximum power output of the scroll machine could be achieved to 10 kW [29], but the power output of the scroll expander beyond 5 kW is barely observed in available literature. On the system scale between 5 and 10 kW, some of expansion machines have been used in ORC system such as axial turbine [30], radial turbine [31], and the single-screw expander [7]. The scroll expander with volume ratio of 5.26 and maximum power output of 10 kW has been commercialized [32], but the tested results are rarely recorded. In addition, a geometric design of the scroll expander for the bigger operating temperature range has been studied [33]. Accordingly, the development of scroll expander for the power output between 5 and 10 kW need to be further studied. There are numerous theoretical studies for optimizing working fluids based on various viewpoints. For instance, Wang et al. [34] considered the screening criteria included heat exchanger area per unit power output and heat recovery efficiency. Lakew et al. [35] screened the working fluids based on power production capability and system components (heat exchanger and turbine size) requirements. Wang et al. [36] pointed out that the working fluids with low Jacob number shows attractive performance for low grade waste heat. The optimal working fluids have been screened out from aforementioned contributions, i.e., R123, R141b, R22, R290, R134a and R227ea. Compare with the experimental publications, the working fluids that commonly used were R123, R245fa or R134a, because the characteristics of the fluids such as chemical stability, corrosion, operating pressure, toxicity and environmental friendly, should be considered. Chang et al. [19] have experimentally tested the performance of the scroll type expanders in parallel with CFD (Computational Fluid Dynamics) simulations for the scroll expanders with different geometrics in an ORC loop. Result showed that the geometry of the tip scroll and the scroll length significantly influence the performance of the expander. The similar calculation results are also presented from Clemente et al. [37]. For a real cycle, the performance of the ORCs vary with several complicate factors, such as the medium of the waste heat source, heat sink condition, the load of the system and the expansion machine. Therefore, precisely controlling the operating conditions of the cycle is important. Accordingly, a cost-effective controller must be developed for the ORC system. Wei et al. [38] established the dynamic model to validate the dynamic behavior of the ORC system, the simulated results presented a very good accuracy. Quoilin et al. [39] simulated three control strategies: (1) a constant evaporating temperature, (2) the evaporating temperature depending on the working conditions, and (3) setting the pump speed according to the expander speed. They found that the third
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Table 1 Some experimental results of the ORC system from available literature (sorted by expander type and power output below 5 kW). Authors Nguyen et al. [16] Li et al. [17] Pei et al. [18] Chang et al. [19] Zhou et al. [20] Quoilin et al. [21] Bracco et al. [22] Declaye et al. [43] Yun et al. [23] Saitoh et al. [24] Manolakos et al. [25] Farrokhi et al. [26] Qiu et al. [3] Zheng et al. [27]
Working fluids n-Pentane R123 R123 R245fa R245fa R123 R245fa R245fa R245fa R113 R134a Isopentane HFE7000 R245fa
Expander type R.T. R.T. R.T. S S S S S S S S M.V. M.V. R.P.
Lubricant No No No Yes Yes No Yes No No No No No No No
gexp 49.8 54–68 65 60.7–80 No data 42–68b 57–74 27–75.7b 44–61.4b 65 30–50 37–45.5 52.4–55.5 40–43
gcycle a
4.3 4.6–8.2 6.8 4.24–7.77 2–8.5 1.7–7.4 7–8.7 0.1–8.54 3.2–7.5 11 3.5–5 2.59–3.1 3.73–3.89 5–6
_ exp W c
1.44 0.7–2.4 1.36 0.53–1.74 0.075–0.645 0.4–1.8b 1.05–1.5c 0.21–2.1b 0.25–3.3b 0.35 0.2–0.95 0.059–0.15 1.66–1.72 0.18–0.35
DTcycle 31–81 3.8–97.6 28.2–101 20.7–88.3 8–140 13.2–163.2d No data 13–97.5 24–120d 35–136 No data 28.3–72.1 5–126.9 23–87.7
R.T.: radial turbine; S: scroll; M.V.: multi-vane; R.P.: rolling piston. a The cycle efficiency is defined as electricity output by the generator divided to the absorbed energy in the evaporator. b The measured power or efficiencies of the expander were obtained from the torque meter. c The electricity delivery by the generator. d The cycle temperature difference is defined as the heat source and heat sink temperature.
strategy did not consistently maintain superheating at the expander inlet, resulting in unsteady operation compared with the first and second strategies. Sun et al. [40] established a mathematical control model and searched for optimal operating strategies to achieve either the best cycle efficiency or net power generation. Zhang et al. [41] incorporated a linear quadratic regulator with a PI controller to simulate the control performance of ORC. Lee et al. [42] experimentally examined the system behaviors, and found that the power output, condensing pressure and evaporating pressure significantly varied with the flow rate of the cooling water in the condenser. To date, very few publications have experimentally investigated the operating parameters such as the subcooling at the pump inlet and the superheating at the expander inlet. The objective of this study firstly describes the characteristic of the scroll expander in the experimental system, presents the test bench of the ORC loop, and the equipment and method for controlling the system. In the second part, the cycle performance operated in fixed superheating at the expander inlet will be presented. In present study, the superheating is defined as the temperature difference between the expander inlet and the corresponding saturated temperature of the inlet pressure. In the third part, the effect of various superheating at the expander inlet with respect to overall performance will be experimentally investigated. In addition, a comparison of the experimental performance between the cycle operated in higher superheating and lower superheating will be discussed. Last part of this paper integrates previous experimental data [19] to compare system performance with different geometries of the scroll expanders. Noted that the system configuration among individual components is unchanged (include measure positions) for the purpose of remaining the consistency of estimating the system performance. This work can support screening an applicable scroll type expander for the heat source temperature below 100 °C. 2. Description of the test rig 2.1. ORC loop In this experimental apparatus, refrigerant HFC-245fa has been utilized as the working fluid. This refrigerant is chemically stable and non-flammable, having a moderate critical temperature for the application in the low-temperature waste heat recovery. This fluid has an appropriate boiling temperature at atmospheric pressure. The refrigerant oil has been mixed with working fluid at a moderate percentage and circulated in the loop. The effect of the
refrigerant oil on the thermo-physical properties of the refrigerant mixture is supposed negligible. Fig. 1 schematically depicts the experimental bench of the ORC, the corresponding main measured state points of the cycle are marked in the figure. The plunger pump was employed to increase the pressure head of the system. The achievable maximum delivered pressure head and flow rate are 50 bars and 18 L/min, respectively. The plunger pump has a high volume ratio, good for operating on small flow rate, and good performance curve. The pump is powered by electrical motor using pulley and belt coupling with a ratio of 1:2, means that as the pump revolves once, the electrical motor rotates twice. The rotational speed of the motor is manually controlled by a frequency inverter. The hot water which was heated by electric heaters is used as a dummy waste heat source. The cooling tower connected to a centrifugal pump delivers cold water into the condenser as a heat sink. Brazed flat plate heat exchanger has a good heat transfer efficiency and compact size. In this study, two heat exchangers with cross-section of 124 mm 504 mm, composed of 80 plates, which was utilized as the evaporator and condenser, respectively. A refrigerant tank is located downstream of the condenser to provide stable supply of working fluid at liquid state. A gear pump is placed between the expander module and the refrigerant tank to pump the refrigerant oil into the circuit. 2.2. Measurement devices In this cycle, temperatures and pressures are measured by T-type thermocouples and piezo-resistive pressure transmitters, respectively. The rotational speed of the expander is simultaneously recorded by a tachometer, and the data are sent through a USB cable to a computer. Turbine flow meter is located downstream of the plunger pump to measure volumetric flow rate of the system. The generated voltage and current from the permanent magnet generator are measured from the data logger and a clamp ammeter respectively. Except those measured by the tachometer, the measured data were recorded by an Agilent 34972A data logger and stored in a computer. Table 2 presents the accuracy and measurement range of the sensors. 2.3. Scroll expander In this study, a scroll-type expander, which was modified from commercial oil-free scroll type air compressor (Anest Iwata SL-165E) with a built-in volume ratio of 4.05 has been employed
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Fig. 1. Schematic diagram of ORC test bench.
Table 2 Accuracy and measured ranges of the sensors. Measure device
Range
Accuracy
Pressures Refrigerant flow rate Temperature Rotational speed
0–30 bar 1.9–37.9 L/min 160 to 400 °C 10–99,999 RPM
±0.5% Full scale ±3% ±0.5 °C ±0.04%
in the ORC system. Using commercial scroll compressor to be an expander is a convenient way for the experimental phase. In this apparatus, the generated shaft work from scroll expander was directly coupled with a three-phase permanent-magnet generator by pulley and belt. The ratio of the rotational speeds between the expander and the generator is 1:1 under all tested conditions. This generator is suitable for wide range of rotational speed and it does not require an auxiliary excitation circuit. It should be noted that the performance of the generator is beyond scope of this study. Table 3 presents basic geometrical information about this scroll type expander. Fig. 2 schematically depicts the scroll expander module. The expander was placed in a cylindrical container to prevent the leakage of working fluid into the environment. Similar treatment on the expander can also be seen in the experimental facilities from Declaye et al. [43]. Since the container has quite large diameter in the experimental apparatus (around
Table 3 Basic geometrical information of the scroll type expander. Basic circle radius (mm) Wrap height (mm) Pitch (mm) Wall thickness (mm) Suction volume (cm3/rev)
3.262 28 20.8 4.5 34.2
46 cm), additional ribs and front flanges are installed to prevent deformation from the working fluid in the container. A small hole with inside diameter of 10.9 mm was drilled at the bottom of the container to enable refrigerant oil to be pumped into the liquid tank. As the high pressure working fluid enters the expander, it is basically a mixture of refrigerant and lubricant during the experimental phase. Once the working fluid leaves the scroll wrap, the refrigerant presents superheated mixture in the expander module, portion of the refrigerant oil will adhere to the expander module (mostly on its back flange) and then flow down to the bottom of the container. As the oil accumulated to a certain level in the sight glass, the gear pump will be manually started for around 3–5 s to pump refrigerant oil to the tank until the sight glass presents empty. It should be reminded that the emptying mechanism was established for this experiment only. Fig. 3 shows the experimental apparatus of the ORC. It should be noted that the cooling tower is not shown in the picture, which was placed outside of the experimental apparatus. 3. Thermodynamic equations The thermo-properties of working fluid on each process in the cycle can be measured by means of two independent properties. In this study, the thermodynamic properties of working fluid are calculated by REFPROP environment [44]. Fig. 1 indicates the main measured locations of the cycle. The mathematical model and cycle performance of each component is described as follows:Thermodynamic pumping power:
_ pp ¼ mðh _ 1 h2 Þ W
ð1Þ
_ denotes the mass flow rate of the working fluid. The pump where m efficiency gpp is defined as pumping work on isentropic process _ ele;motor : divided by electricity consumed by the motor, W
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Fig. 2. Schematic diagram of expander module.
4. Results and discussions 4.1. System operated with fixed superheating at expander inlet
ð2Þ
In each experiment, at least 20 min of steady-state were maintained for data recording to evaluate system performance. Each operating condition, such as system pressure difference, rotational speed of the expander, and the superheating at the inlet of the expander will be primarily controlled by heat source temperature, the system flow rate and the load of the generator. In this section, the superheating at the expander inlet was controlled at 3 ± 1 K, for the purpose of preventing the potential measurement oscillation and measurement uncertainty of the sensors. Table 4 presents the main operating ranges of the experiment. According to the theory of error propagation, the measurement uncertainties are calculated using the root-sum-square method. The uncertainty UY of the variable Y is calculated as a function of the uncertainties U xi for each measured variable xi, which is expressed in Eq. (8) [45,46].
ð3Þ
vffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi u 2 uX @Y UY ¼ t U 2xi @x i i
Fig. 3. Photograph of the ORC apparatus.
_ W
gpp ¼ _ pp;is W ele;motor Evaporating process:
_ 3 h2 Þ Q_ ev ap ¼ mðh Expansion work from scroll expander is calculated by:
_ exp ¼ mðh _ 3 h4 Þ W
ð4Þ
ð8Þ
As shown in Fig. 4, the cycle efficiency mainly varied with the rotational speed of the expander and pressure difference of the system. Slight deviation of rotational speed is due to the load of
The isentropic efficiency of the expander is expressed as:
gexp ¼
h3 h4 h3 h4;is
ð5Þ
The residual heat is removed from the condenser, Q_ cond is written as:
_ 1 h4 Þ Q_ cond ¼ mðh
ð6Þ
Cycle efficiency:
gcycle ¼
h1 h2 þ h3 h4 h3 h2
ð7Þ
Table 4 Operating conditions of the experiment. Operating conditions
Minimum value
Maximum value
Hot water inlet temperature Hot water outlet temperature Cold water inlet temperature Cold water outlet temperature Refrigerant flow rate Expander rotational speed Expander inlet pressure Pump inlet pressure
77.9 °C 72.5 °C 11.5 °C 15.6 °C 2.1 L/min 1535 RPM 6.54 bar 1.43 bar
95.6 °C 88.6 °C 20.1 °C 24.2 °C 4.74 L/min 2970 RPM 9.47 bar 1.82 bar
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Fig. 4. The influence of system pressure range and rotational speeds of the expander on ORC efficiency.
resistors cannot be precisely regulated to a specific rotational speed. As increasing load of resistors, the rotational speed of the expander decreased but the cycle efficiency increased. Meanwhile, the electric current generated from the generator simultaneously increased and the voltage output decreased because higher electric current increasing temperature of the coil, and therefore reduced magnet intensity of the generator. The maximum system efficiency of 9.44% is achieved at a rotational speed of 1535 RPM whereas minimum system efficiency of 4.87% is presented at 2970 RPM. It can be seen that the system is preferred to operate at lower rotational speed in order to maintain favorable efficiency. Fig. 5 shows expander efficiency as a function of the inlet temperature of the expander. The maximum expander efficiency of 73.1% is achieved on 1535 RPM and an inlet temperature of the expander is 90.6 °C. As rotational speed operated between 1535 RPM and 2130 RPM, the efficiency presented a nearly flat trend over tested results. As rotational speed beyond 2130 RPM, the isentropic efficiency dramatically decreases; it may be due to radial and axial leakage, the pressure drop during suction processes, friction losses between fixed and orbiting scrolls, and losses in the bearings. As described above, some loss factors have been evaluated by semi-empirical approach [45,47]. It should be noted that the measurement of the above losses of the expander are very
Fig. 5. Expander efficiency as a function of expander inlet temperature.
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difficult in this experimental apparatus. Therefore, this study does not focus the quantification on above individual losses. Since the cycle performance is significantly influenced by the rotational speed of the expander and the pressure difference of the system, the generated power from the expander simultaneously vary with the above factors. Fig. 6 plots the cycle efficiency and power output as a function of the rotational speed of the expander. It should be noted that the solid symbols represent cycle efficiency and the hollow symbols represent the power output of the expander. As can be seen, the power output slowly increases with rotational speed because the refrigerant flow rate of the system is simultaneously increased, whereas the cycle efficiency presents an opposite trends. The main reason is the fact that of the efficiency of the expander obviously decreases as rotational speed increases. For this reason, as the cycle is operated at a higher pressure difference (such as 8 bars in this experiment), a compromise operating point appears at a lower rotational speed (about 2090 RPM). As pressure difference is operated at 5 bars, the compromise operating point approximately locates on 3028 RPM. From Fig. 6, the maximum power produced by scroll expander of 2.3 kW is observed, the corresponding recorded rotational speed of 2970 RPM and the pressure difference of 8 bars. It can be known that as the expander is operated at a fixed rotational speed and superheating, the power output increases with increasing of the system pressure difference, because it implies a higher inlet pressure and density, which involves higher mass flow rate in the expander. In the above operating condition, the maximum electricity produced from the generator only 1.56 kW as pressure difference and rotational speed is 8 bars and 2970 RPM. All the experiment data in this manuscript were recorded under steady-state operations. Fig. 7 shows a sample run, which demonstrates steady-state recorded data when the cycle was operated at a rotational speed of 1535 ± 40 RPM and a cycle pressure difference of 8 bars. In this duration, the hot water inlet and cold water inlet were maintained at an average temperature of 94.6 °C and 11.5 °C, respectively. As can be seen, the experimental results are highly stable over a time period of 20 min. As shown in Fig. 8, the pump efficiency increased with increasing of pressure difference of the system. The maximum pump efficiency and system pressure difference is 33.6% and 8 bar, the corresponding power consumption by the motor was 0.204 kW. Increasing rotational speed of the expander involves more refrigerant delivery, and improves cooling performance of the plunger
Fig. 6. Influence of rotational speed to the cycle efficiency and power output. The solid symbols represent the cycle efficiency, and hollow symbols represent the power output of the expander.
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Fig. 7. Steady-state data at cycle pressure difference of 8 bars and rotational speed of expander is 1535 ± 40 RPM.
Fig. 8. Pump efficiency as a function of pressure difference of the system and rotational speed of the expander.
pump at an operating pressure difference of the system. In this experimental apparatus, the pump efficiency is too low, and requires large subcooling (between 11 and 13 °C) at inlet to ensure stable operation. As the system pressure difference and rotational speed are increased, the heat rejection of ORC is simultaneously increased. This is primarily due to the increase tendency of the cycle efficiency is low as operating pressure difference in the experiment is increased. As shown in Fig. 9, the maximum heat rejection of the condenser was 24.8 kW at a rotational speed of 2970 RPM and a system pressure difference of 8 bars. In which, the outlet temperature of the expander was 53.6 °C, whereas the minimum recorded heat rejection was 10.79 kW. As shown in Fig. 9, it is observed that the tendency of the heat rejection is more precipitous as the cycle is operated at a higher rotational speed of the expander.
Fig. 9. System heat rejection versus system pressure difference.
experimental data concerning the various superheating at the expander inlet are described below. Fig. 10 plots the influence of superheating versus the expander and ORC efficiency, respectively. One must be reminded that the solid and hollow symbols represent expander efficiency and ORC efficiency, respectively. It should be noted that the expander was controlled at rotational speed of 2456 ± 50 RPM, and the system pressure difference was operated at 5 bar and 6 bar. These data were obtained by carefully regulating the flow rate and the load of the system (i.e., the resistors). As seen in the figure, the expander efficiency increases with increasing superheating level, because the density and the collision frequencies among molecules of the working fluid is reasonably reduced as increasing the superheating. As a result, the entropy generation during the expansion process is simultaneously reduced. Similar tendency can be found in the publication of Bracco et al. [22]: as the superheating was operated between 10 and 25 °C and at fixed rotational speed of the pump, the expander efficiency, cycle efficiency and power output simultaneously increased, but this trend became insignificant as the superheating was higher than 40 °C in a fixed expansion ratio from their experiment. In this experiment, the maximum efficiency of the expander is 76.8% at a system pressure difference of 5 bars and 72.2% at 6 bars. Similarly, the ORC efficiency increases with the superheating of the expander. The maximum cycle efficiency
4.2. Effect of superheating at expander inlet on the system performance This section discusses the influence of superheating of the expander versus the overall performance of the system. The cycle was fed with hot water at a pressure of one atmosphere. The
Fig. 10. Effect of superheating at expander inlet on expander and cycle performance. The solid symbols represent expander efficiency, and hollow symbols represent the cycle efficiency.
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of 8.38% was obtained with superheating at 14.2 °C and a cycle pressure difference of 6 bar, whereas 8.12% was obtained with 19.7 °C and 5 bar. During the experimental phase, the efficiency of the pump did not obviously vary with the superheating because of the pump performance is mainly a function of the operating pressure difference. Two experimental results are compared in Fig. 11, which based on the fixed pressure difference of 6 bars, and operated at the superheating of 3.3 °C and 14.2 °C, respectively. This figure reveals that as the cycle is operated at higher superheating, the pinch point temperature difference of the evaporator is higher than that of lower superheating, whereas the deviation of the pinch point temperature difference of the condenser is very small as the cycle is operated at different superheating. One must be reminded that the system was imposed higher refrigerant flow rate as the cycle was operated at a lower superheating. The record of the mass flow rate of 0.075 kg/s and 0.0708 kg/s was obtained as the superheating was operated at 3.3 °C and 14.2 °C, respectively. Fig. 11 shows that as the superheating is operated at 3.3 °C, the temperature difference between hot water inlet and expander inlet is higher 1.2 °C than operated at 14.2 °C. As the cycle is operated at the superheating of 14.2 °C, the specific irreversibility of the evaporator is bigger than that of 3.3 °C, the ‘‘redundant specific irreversibility’’ is depicted in the gray area as the superheating were operated at 14.2 °C and 3.3 °C, respectively. In the other words, the specific irreversibility of the evaporator during the heat exchange process is relatively higher, if the cycle is operated at higher superheating. On the other hand, the difference of the specific irreversibility of the condenser is not significant as the cycle operated at the superheating between 14.2 °C and 3.3 °C.
4.3. The integration of experimental performance for different built-in volume ratio of scroll expanders The integrated experimental performance with respect to different scroll-type expanders are shown in Fig. 12(a), the
Fig. 12. Experimental performance based on different built-in volume ratio of the scroll expander (a) expander power output versus cycle efficiency; (b) heat source temperature versus operating pressure difference of the cycle.
Fig. 11. T-s diagram of the cycle operated with superheating at 3.3 °C (red) and 14.2 °C (blue). The red cycle represents the expander operated at 2384 RPM, and blue cycle represents the expander operated at 2433 RPM. (For interpretation of the references to color in this figure legend, the reader is referred to the web version of this article.)
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corresponding heat source temperature and operating pressure difference of the system in each experiment are shown in Fig. 12(b). In this figure, symbols Vr2 and Vr3 express the different built-in volume ratio of the scroll type expander obtained by previous test [19]. In previous experiment, the cycle was also fed with hot water at a pressure of one atmosphere as the heat transfer fluid. The previous experimental results involve rotational speed between 1641 RPM and 3054 RPM of the expander and a fixed superheating of 3 ± 1 K. Fig. 12 conveys that as the cycle is operated at the higher pressure difference, the system presents predominant performance as the scroll expander with bigger built-in volume ratio is utilized. In view of this, the expander used in the present study is inadequate to operate at a pressure difference of 5 bars in the system, because the output power of the expander is reduced to prohibitive value. Based on the integrated experimental data, the expander used in present study is preferred to operate at the pressure difference higher than 6 bars. 5. Conclusions This study experimentally investigates the performance of a scroll expander with a built-in volume ratio of 4.05 in an ORC system using HFC-245fa as working fluid, and the fluid has been mixed with moderate refrigerant oil that circulated in the system. The expander was modified from an oil-free open-drive scroll type air compressor and operated in reverse mode. The hot water at a pressure of one atmosphere was used as a dummy waste heat source and a cooling tower served as a heat sink. The main findings are as follows: 1. As the cycle is operated in fixed superheating of 3 ± 1 K, the maximum cycle efficiency is 9.43% at a rotational speed of 1535 ± 40 RPM and a cycle pressure difference of 8 bars. The maximum power of the expander is 2.3 kW at a rotational speed of 2970 ± 50 RPM and an operating cycle pressure difference of 8 bars. 2. The expander efficiency decreases as the rotational speed increases. A maximum expander isentropic efficiency is 73% at a rotational speed of 1535 ± 40 RPM and a cycle pressure difference of 8 bars. As the expander is operated between 1535 and 2130 RPM, the isentropic efficiency almost remains identical under above rotational speeds. 3. The maximum power produced by the expander is not corresponding to the maximum cycle efficiency for the real cycle. The compromise operating conditions have relation with system pressure difference and rotational speed of the expander. 4. Both efficiencies of the expander and the cycle increase with the increasing of superheating, but the irreversibility of the evaporator also increase simultaneously. 5. Experimental data from the scroll type expanders with different built-in volume ratio have been integrated. The expander used in this study is preferred to operate at higher range of pressure difference.
Acknowledgements The authors would like to appreciate the financial supports provided by Ministry of Science and Technology of the Republic of China, Taiwan (Contract No. MOST 103-2221-E-027-106); National Taipei University of Technology, Taipei, Taiwan (Contract No. 103T140); especially appreciate Kwan Chiu Radio Manufacturing Co., Ltd. for their experimental place and technical supports.
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Experimental study on low-temperature organic Rankine cycle utilizing scroll type expander Jen-Chieh Chang a, Tzu-Chen Hung b,⇑, Ya-Ling He c, Wenping Zhang d a
Graduate Institute of Mechanical and Electrical Engineering, National Taipei University of Technology, Taipei, Taiwan Department of Mechanical Engineering, National Taipei University of Technology, Taipei, Taiwan c Key Laboratory of Thermo-Fluid Science and Engineering of MOE, School of Energy and Power Engineering, Xi’an Jiaotong University, Xi’an, Shaanxi, PR China d College of Power and Energy Engineering, Harbin Engineering University, Harbin, Heilongjiang Province 150001, PR China b
h i g h l i g h t s Experimental studies of the ORC system operating with R245fa are presented. Scroll type expander was employed in the ORC system. The cycle was fueled with hot water (maximum 95.6 °C). Superheating improves isentropic efficiency of the expander.
a r t i c l e
i n f o
Article history: Received 6 February 2015 Received in revised form 11 May 2015 Accepted 25 May 2015
Keywords: Organic Rankine cycle Low-temperature Scroll expander Waste heat recovery Superheating
a b s t r a c t This paper focuses on experimental performance of an open-drive scroll type expander in an organic Rankine cycle (ORC) system. The expander was an originally oil-free scroll type air compressor with a built-in volume ratio of 4.05. The cycle used HFC-245fa as working fluid, and the loop has been mixed with a moderate concentration of refrigerant oil that circulated in the cycle. The experimental results of this study are divided into two main parts: first part focuses the experimental performance on the fixed superheating at the expander inlet with respect to various pressure differences of the system and rotational speeds of the expander. Second part involves various superheating at the expander inlet which was operated at fixed rotational speed and operating pressure difference of 5 bars and 6 bars. When the cycle was operated under fixed superheating conditions, the maximum cycle efficiency, expander efficiency and power output of the expander are 9.44%, 73.1% and 2.3 kW respectively. On the other hand, when the expander is operated in various superheating conditions, the expander and cycle efficiency simultaneously increase with the increasing of superheating. In addition, this paper not only focuses on the experimental results using the current expander, but also integrates the previous experimental data with present study to identify an appropriate scroll type expander with respect to various operating pressure differences for the heat source below 100 °C. Ó 2015 Elsevier Ltd. All rights reserved.
1. Introduction Organic Rankine cycle (ORC) is a promising machine for energy conversion from low-temperature heat sources. U.S. Department of Energy (U.S. DOE) has indicated that approximately 60% of low-temperature waste heat (below 232 °C) is not recovered, and this heat is in large quantities [1]. In order to convert low-temperature heat sources to useful work, Yamamoto et al. [2] experimentally demonstrated that the
⇑ Corresponding author. Tel.: +886 2 2771 2171x2021; fax: +886 2 2731 7191. E-mail address: [email protected] (T.-C. Hung). http://dx.doi.org/10.1016/j.apenergy.2015.05.118 0306-2619/Ó 2015 Elsevier Ltd. All rights reserved.
use of a low boiling point working fluid, such as R-123 utilized in the Rankine cycle, result in a better cycle performance than using water for the turbine inlet temperature below 120 °C. For this reason, the ORC systems have been widely tested for various applications. Qiu et al. [3] tested a biomass-fired CHP (combined heat and power) with an ORC for domestic application. Yagoub et al. [4] tested the ORC system driven by solar and natural gas in an office building, it reported that the system can effectively save the electricity and heat demand of the building. Wang et al. [5] tested both evacuated solar collectors and the plate collectors for absorbing solar energy in an ORC system, the result shows that the overall
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Nomenclature h _ m Q_
DP P s T _ W
enthalpy (kJ/kg) mass flow rate (kg/s) thermal power (kW) pressure difference of the system, P3 P1 (bar) pressure (bar) specific entropy (kJ/kg K) temperature (°C) work (kW)
Greek symbols g efficiency
thermodynamic power generation efficiency is 4.2%, when the evacuated solar collectors is utilized in the ORC, and with the condition of plat solar collector, it is about 3.2%. Wang et al. [6] tested a flat-plate collector for gathering solar energy in a regenerative ORC, and the result shows that both thermal and collector efficiencies could be improved significantly by adjusting the working fluid flow rate to an appropriate level. Zheng et al. [7] experimentally tested an ORC system for recovering the exhaust gas from a diesel engine, result shows that the maximum improvement of overall system efficiency was 1.53%. Endo et al. [8] incorporated an ORC in the automotive engine, they reported that the thermal efficiencies can be improved about 10%. ORCs also have been investigated to find a maximum work output from various combination thermodynamic cycles, such as recovering exhaust heat from the gas turbine [9], the high-temperature gas-cooled reactors [10], and the large marine diesel engine applications [11]. The applications of ORC system have a wide spectrum of recovering various waste heat sources or renewable energy. The key factor of the ORC is the selection of working fluids. Hung et al. [12,13] reported that the system efficiency increases and decreases for wet and dry fluids, respectively, and the isentropic fluids achieve an approximately constant value as increase the superheating of the expander inlet. Furthermore, the theoretical report has shown that wet fluids with very steep saturated vapor curves in T-s diagram have a better overall performance in energy conversion efficiencies than that of dry fluids [14]. ORCs are commonly operated at relatively low pressure differences in finite temperature ranges. Therefore, it is very important to reduce losses of the components of the system in order to achieve higher performance, especially in small-scale ORC systems. In the expansion machine, Bao et al. [15] summarized the advantages and disadvantages of various type expanders used in the ORC system. They indicated that the scroll and multi-vane expander are applicable to the power output smaller than 10 kW for the ORCs. Nevertheless, numerous publications mainly focused on ORCs with power output smaller than 5 kW in the laboratory tested. Table 1 summarizes recently experimental results which focus on the power output below 5 kW, without utilizing regenerator in the ORC system. It seems like that the positive-displacement expander such as scroll type expander is widely used for this scale of ORCs, because of the scroll machine can tolerate two-phase conditions [28], compact size, low cost, high efficiency and lower moving-parts. Bao et al. [15] reported that the capacity of the multi-vane expander could be reached to 10 kW. However, according to available literature, the maximum power output was only 1.72 kW [3]. Multi-vane expander has a simple structure, easier to manufacture and low cost, but shows a relatively low efficiency over positive-displacement expanders
Subscript cond ele exp evap H L in is out pp
condenser electricity expander evaporator hot, high low inlet isentropic outlet pump
(refer to Table 1), it might be a main reason why the experimental studies on these expansion machine are very few in the literature. Recently, it was claimed that the maximum power output of the scroll machine could be achieved to 10 kW [29], but the power output of the scroll expander beyond 5 kW is barely observed in available literature. On the system scale between 5 and 10 kW, some of expansion machines have been used in ORC system such as axial turbine [30], radial turbine [31], and the single-screw expander [7]. The scroll expander with volume ratio of 5.26 and maximum power output of 10 kW has been commercialized [32], but the tested results are rarely recorded. In addition, a geometric design of the scroll expander for the bigger operating temperature range has been studied [33]. Accordingly, the development of scroll expander for the power output between 5 and 10 kW need to be further studied. There are numerous theoretical studies for optimizing working fluids based on various viewpoints. For instance, Wang et al. [34] considered the screening criteria included heat exchanger area per unit power output and heat recovery efficiency. Lakew et al. [35] screened the working fluids based on power production capability and system components (heat exchanger and turbine size) requirements. Wang et al. [36] pointed out that the working fluids with low Jacob number shows attractive performance for low grade waste heat. The optimal working fluids have been screened out from aforementioned contributions, i.e., R123, R141b, R22, R290, R134a and R227ea. Compare with the experimental publications, the working fluids that commonly used were R123, R245fa or R134a, because the characteristics of the fluids such as chemical stability, corrosion, operating pressure, toxicity and environmental friendly, should be considered. Chang et al. [19] have experimentally tested the performance of the scroll type expanders in parallel with CFD (Computational Fluid Dynamics) simulations for the scroll expanders with different geometrics in an ORC loop. Result showed that the geometry of the tip scroll and the scroll length significantly influence the performance of the expander. The similar calculation results are also presented from Clemente et al. [37]. For a real cycle, the performance of the ORCs vary with several complicate factors, such as the medium of the waste heat source, heat sink condition, the load of the system and the expansion machine. Therefore, precisely controlling the operating conditions of the cycle is important. Accordingly, a cost-effective controller must be developed for the ORC system. Wei et al. [38] established the dynamic model to validate the dynamic behavior of the ORC system, the simulated results presented a very good accuracy. Quoilin et al. [39] simulated three control strategies: (1) a constant evaporating temperature, (2) the evaporating temperature depending on the working conditions, and (3) setting the pump speed according to the expander speed. They found that the third
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Table 1 Some experimental results of the ORC system from available literature (sorted by expander type and power output below 5 kW). Authors Nguyen et al. [16] Li et al. [17] Pei et al. [18] Chang et al. [19] Zhou et al. [20] Quoilin et al. [21] Bracco et al. [22] Declaye et al. [43] Yun et al. [23] Saitoh et al. [24] Manolakos et al. [25] Farrokhi et al. [26] Qiu et al. [3] Zheng et al. [27]
Working fluids n-Pentane R123 R123 R245fa R245fa R123 R245fa R245fa R245fa R113 R134a Isopentane HFE7000 R245fa
Expander type R.T. R.T. R.T. S S S S S S S S M.V. M.V. R.P.
Lubricant No No No Yes Yes No Yes No No No No No No No
gexp 49.8 54–68 65 60.7–80 No data 42–68b 57–74 27–75.7b 44–61.4b 65 30–50 37–45.5 52.4–55.5 40–43
gcycle a
4.3 4.6–8.2 6.8 4.24–7.77 2–8.5 1.7–7.4 7–8.7 0.1–8.54 3.2–7.5 11 3.5–5 2.59–3.1 3.73–3.89 5–6
_ exp W c
1.44 0.7–2.4 1.36 0.53–1.74 0.075–0.645 0.4–1.8b 1.05–1.5c 0.21–2.1b 0.25–3.3b 0.35 0.2–0.95 0.059–0.15 1.66–1.72 0.18–0.35
DTcycle 31–81 3.8–97.6 28.2–101 20.7–88.3 8–140 13.2–163.2d No data 13–97.5 24–120d 35–136 No data 28.3–72.1 5–126.9 23–87.7
R.T.: radial turbine; S: scroll; M.V.: multi-vane; R.P.: rolling piston. a The cycle efficiency is defined as electricity output by the generator divided to the absorbed energy in the evaporator. b The measured power or efficiencies of the expander were obtained from the torque meter. c The electricity delivery by the generator. d The cycle temperature difference is defined as the heat source and heat sink temperature.
strategy did not consistently maintain superheating at the expander inlet, resulting in unsteady operation compared with the first and second strategies. Sun et al. [40] established a mathematical control model and searched for optimal operating strategies to achieve either the best cycle efficiency or net power generation. Zhang et al. [41] incorporated a linear quadratic regulator with a PI controller to simulate the control performance of ORC. Lee et al. [42] experimentally examined the system behaviors, and found that the power output, condensing pressure and evaporating pressure significantly varied with the flow rate of the cooling water in the condenser. To date, very few publications have experimentally investigated the operating parameters such as the subcooling at the pump inlet and the superheating at the expander inlet. The objective of this study firstly describes the characteristic of the scroll expander in the experimental system, presents the test bench of the ORC loop, and the equipment and method for controlling the system. In the second part, the cycle performance operated in fixed superheating at the expander inlet will be presented. In present study, the superheating is defined as the temperature difference between the expander inlet and the corresponding saturated temperature of the inlet pressure. In the third part, the effect of various superheating at the expander inlet with respect to overall performance will be experimentally investigated. In addition, a comparison of the experimental performance between the cycle operated in higher superheating and lower superheating will be discussed. Last part of this paper integrates previous experimental data [19] to compare system performance with different geometries of the scroll expanders. Noted that the system configuration among individual components is unchanged (include measure positions) for the purpose of remaining the consistency of estimating the system performance. This work can support screening an applicable scroll type expander for the heat source temperature below 100 °C. 2. Description of the test rig 2.1. ORC loop In this experimental apparatus, refrigerant HFC-245fa has been utilized as the working fluid. This refrigerant is chemically stable and non-flammable, having a moderate critical temperature for the application in the low-temperature waste heat recovery. This fluid has an appropriate boiling temperature at atmospheric pressure. The refrigerant oil has been mixed with working fluid at a moderate percentage and circulated in the loop. The effect of the
refrigerant oil on the thermo-physical properties of the refrigerant mixture is supposed negligible. Fig. 1 schematically depicts the experimental bench of the ORC, the corresponding main measured state points of the cycle are marked in the figure. The plunger pump was employed to increase the pressure head of the system. The achievable maximum delivered pressure head and flow rate are 50 bars and 18 L/min, respectively. The plunger pump has a high volume ratio, good for operating on small flow rate, and good performance curve. The pump is powered by electrical motor using pulley and belt coupling with a ratio of 1:2, means that as the pump revolves once, the electrical motor rotates twice. The rotational speed of the motor is manually controlled by a frequency inverter. The hot water which was heated by electric heaters is used as a dummy waste heat source. The cooling tower connected to a centrifugal pump delivers cold water into the condenser as a heat sink. Brazed flat plate heat exchanger has a good heat transfer efficiency and compact size. In this study, two heat exchangers with cross-section of 124 mm 504 mm, composed of 80 plates, which was utilized as the evaporator and condenser, respectively. A refrigerant tank is located downstream of the condenser to provide stable supply of working fluid at liquid state. A gear pump is placed between the expander module and the refrigerant tank to pump the refrigerant oil into the circuit. 2.2. Measurement devices In this cycle, temperatures and pressures are measured by T-type thermocouples and piezo-resistive pressure transmitters, respectively. The rotational speed of the expander is simultaneously recorded by a tachometer, and the data are sent through a USB cable to a computer. Turbine flow meter is located downstream of the plunger pump to measure volumetric flow rate of the system. The generated voltage and current from the permanent magnet generator are measured from the data logger and a clamp ammeter respectively. Except those measured by the tachometer, the measured data were recorded by an Agilent 34972A data logger and stored in a computer. Table 2 presents the accuracy and measurement range of the sensors. 2.3. Scroll expander In this study, a scroll-type expander, which was modified from commercial oil-free scroll type air compressor (Anest Iwata SL-165E) with a built-in volume ratio of 4.05 has been employed
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Fig. 1. Schematic diagram of ORC test bench.
Table 2 Accuracy and measured ranges of the sensors. Measure device
Range
Accuracy
Pressures Refrigerant flow rate Temperature Rotational speed
0–30 bar 1.9–37.9 L/min 160 to 400 °C 10–99,999 RPM
±0.5% Full scale ±3% ±0.5 °C ±0.04%
in the ORC system. Using commercial scroll compressor to be an expander is a convenient way for the experimental phase. In this apparatus, the generated shaft work from scroll expander was directly coupled with a three-phase permanent-magnet generator by pulley and belt. The ratio of the rotational speeds between the expander and the generator is 1:1 under all tested conditions. This generator is suitable for wide range of rotational speed and it does not require an auxiliary excitation circuit. It should be noted that the performance of the generator is beyond scope of this study. Table 3 presents basic geometrical information about this scroll type expander. Fig. 2 schematically depicts the scroll expander module. The expander was placed in a cylindrical container to prevent the leakage of working fluid into the environment. Similar treatment on the expander can also be seen in the experimental facilities from Declaye et al. [43]. Since the container has quite large diameter in the experimental apparatus (around
Table 3 Basic geometrical information of the scroll type expander. Basic circle radius (mm) Wrap height (mm) Pitch (mm) Wall thickness (mm) Suction volume (cm3/rev)
3.262 28 20.8 4.5 34.2
46 cm), additional ribs and front flanges are installed to prevent deformation from the working fluid in the container. A small hole with inside diameter of 10.9 mm was drilled at the bottom of the container to enable refrigerant oil to be pumped into the liquid tank. As the high pressure working fluid enters the expander, it is basically a mixture of refrigerant and lubricant during the experimental phase. Once the working fluid leaves the scroll wrap, the refrigerant presents superheated mixture in the expander module, portion of the refrigerant oil will adhere to the expander module (mostly on its back flange) and then flow down to the bottom of the container. As the oil accumulated to a certain level in the sight glass, the gear pump will be manually started for around 3–5 s to pump refrigerant oil to the tank until the sight glass presents empty. It should be reminded that the emptying mechanism was established for this experiment only. Fig. 3 shows the experimental apparatus of the ORC. It should be noted that the cooling tower is not shown in the picture, which was placed outside of the experimental apparatus. 3. Thermodynamic equations The thermo-properties of working fluid on each process in the cycle can be measured by means of two independent properties. In this study, the thermodynamic properties of working fluid are calculated by REFPROP environment [44]. Fig. 1 indicates the main measured locations of the cycle. The mathematical model and cycle performance of each component is described as follows:Thermodynamic pumping power:
_ pp ¼ mðh _ 1 h2 Þ W
ð1Þ
_ denotes the mass flow rate of the working fluid. The pump where m efficiency gpp is defined as pumping work on isentropic process _ ele;motor : divided by electricity consumed by the motor, W
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Fig. 2. Schematic diagram of expander module.
4. Results and discussions 4.1. System operated with fixed superheating at expander inlet
ð2Þ
In each experiment, at least 20 min of steady-state were maintained for data recording to evaluate system performance. Each operating condition, such as system pressure difference, rotational speed of the expander, and the superheating at the inlet of the expander will be primarily controlled by heat source temperature, the system flow rate and the load of the generator. In this section, the superheating at the expander inlet was controlled at 3 ± 1 K, for the purpose of preventing the potential measurement oscillation and measurement uncertainty of the sensors. Table 4 presents the main operating ranges of the experiment. According to the theory of error propagation, the measurement uncertainties are calculated using the root-sum-square method. The uncertainty UY of the variable Y is calculated as a function of the uncertainties U xi for each measured variable xi, which is expressed in Eq. (8) [45,46].
ð3Þ
vffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi u 2 uX @Y UY ¼ t U 2xi @x i i
Fig. 3. Photograph of the ORC apparatus.
_ W
gpp ¼ _ pp;is W ele;motor Evaporating process:
_ 3 h2 Þ Q_ ev ap ¼ mðh Expansion work from scroll expander is calculated by:
_ exp ¼ mðh _ 3 h4 Þ W
ð4Þ
ð8Þ
As shown in Fig. 4, the cycle efficiency mainly varied with the rotational speed of the expander and pressure difference of the system. Slight deviation of rotational speed is due to the load of
The isentropic efficiency of the expander is expressed as:
gexp ¼
h3 h4 h3 h4;is
ð5Þ
The residual heat is removed from the condenser, Q_ cond is written as:
_ 1 h4 Þ Q_ cond ¼ mðh
ð6Þ
Cycle efficiency:
gcycle ¼
h1 h2 þ h3 h4 h3 h2
ð7Þ
Table 4 Operating conditions of the experiment. Operating conditions
Minimum value
Maximum value
Hot water inlet temperature Hot water outlet temperature Cold water inlet temperature Cold water outlet temperature Refrigerant flow rate Expander rotational speed Expander inlet pressure Pump inlet pressure
77.9 °C 72.5 °C 11.5 °C 15.6 °C 2.1 L/min 1535 RPM 6.54 bar 1.43 bar
95.6 °C 88.6 °C 20.1 °C 24.2 °C 4.74 L/min 2970 RPM 9.47 bar 1.82 bar
J.-C. Chang et al. / Applied Energy 155 (2015) 150–159
Fig. 4. The influence of system pressure range and rotational speeds of the expander on ORC efficiency.
resistors cannot be precisely regulated to a specific rotational speed. As increasing load of resistors, the rotational speed of the expander decreased but the cycle efficiency increased. Meanwhile, the electric current generated from the generator simultaneously increased and the voltage output decreased because higher electric current increasing temperature of the coil, and therefore reduced magnet intensity of the generator. The maximum system efficiency of 9.44% is achieved at a rotational speed of 1535 RPM whereas minimum system efficiency of 4.87% is presented at 2970 RPM. It can be seen that the system is preferred to operate at lower rotational speed in order to maintain favorable efficiency. Fig. 5 shows expander efficiency as a function of the inlet temperature of the expander. The maximum expander efficiency of 73.1% is achieved on 1535 RPM and an inlet temperature of the expander is 90.6 °C. As rotational speed operated between 1535 RPM and 2130 RPM, the efficiency presented a nearly flat trend over tested results. As rotational speed beyond 2130 RPM, the isentropic efficiency dramatically decreases; it may be due to radial and axial leakage, the pressure drop during suction processes, friction losses between fixed and orbiting scrolls, and losses in the bearings. As described above, some loss factors have been evaluated by semi-empirical approach [45,47]. It should be noted that the measurement of the above losses of the expander are very
Fig. 5. Expander efficiency as a function of expander inlet temperature.
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difficult in this experimental apparatus. Therefore, this study does not focus the quantification on above individual losses. Since the cycle performance is significantly influenced by the rotational speed of the expander and the pressure difference of the system, the generated power from the expander simultaneously vary with the above factors. Fig. 6 plots the cycle efficiency and power output as a function of the rotational speed of the expander. It should be noted that the solid symbols represent cycle efficiency and the hollow symbols represent the power output of the expander. As can be seen, the power output slowly increases with rotational speed because the refrigerant flow rate of the system is simultaneously increased, whereas the cycle efficiency presents an opposite trends. The main reason is the fact that of the efficiency of the expander obviously decreases as rotational speed increases. For this reason, as the cycle is operated at a higher pressure difference (such as 8 bars in this experiment), a compromise operating point appears at a lower rotational speed (about 2090 RPM). As pressure difference is operated at 5 bars, the compromise operating point approximately locates on 3028 RPM. From Fig. 6, the maximum power produced by scroll expander of 2.3 kW is observed, the corresponding recorded rotational speed of 2970 RPM and the pressure difference of 8 bars. It can be known that as the expander is operated at a fixed rotational speed and superheating, the power output increases with increasing of the system pressure difference, because it implies a higher inlet pressure and density, which involves higher mass flow rate in the expander. In the above operating condition, the maximum electricity produced from the generator only 1.56 kW as pressure difference and rotational speed is 8 bars and 2970 RPM. All the experiment data in this manuscript were recorded under steady-state operations. Fig. 7 shows a sample run, which demonstrates steady-state recorded data when the cycle was operated at a rotational speed of 1535 ± 40 RPM and a cycle pressure difference of 8 bars. In this duration, the hot water inlet and cold water inlet were maintained at an average temperature of 94.6 °C and 11.5 °C, respectively. As can be seen, the experimental results are highly stable over a time period of 20 min. As shown in Fig. 8, the pump efficiency increased with increasing of pressure difference of the system. The maximum pump efficiency and system pressure difference is 33.6% and 8 bar, the corresponding power consumption by the motor was 0.204 kW. Increasing rotational speed of the expander involves more refrigerant delivery, and improves cooling performance of the plunger
Fig. 6. Influence of rotational speed to the cycle efficiency and power output. The solid symbols represent the cycle efficiency, and hollow symbols represent the power output of the expander.
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Fig. 7. Steady-state data at cycle pressure difference of 8 bars and rotational speed of expander is 1535 ± 40 RPM.
Fig. 8. Pump efficiency as a function of pressure difference of the system and rotational speed of the expander.
pump at an operating pressure difference of the system. In this experimental apparatus, the pump efficiency is too low, and requires large subcooling (between 11 and 13 °C) at inlet to ensure stable operation. As the system pressure difference and rotational speed are increased, the heat rejection of ORC is simultaneously increased. This is primarily due to the increase tendency of the cycle efficiency is low as operating pressure difference in the experiment is increased. As shown in Fig. 9, the maximum heat rejection of the condenser was 24.8 kW at a rotational speed of 2970 RPM and a system pressure difference of 8 bars. In which, the outlet temperature of the expander was 53.6 °C, whereas the minimum recorded heat rejection was 10.79 kW. As shown in Fig. 9, it is observed that the tendency of the heat rejection is more precipitous as the cycle is operated at a higher rotational speed of the expander.
Fig. 9. System heat rejection versus system pressure difference.
experimental data concerning the various superheating at the expander inlet are described below. Fig. 10 plots the influence of superheating versus the expander and ORC efficiency, respectively. One must be reminded that the solid and hollow symbols represent expander efficiency and ORC efficiency, respectively. It should be noted that the expander was controlled at rotational speed of 2456 ± 50 RPM, and the system pressure difference was operated at 5 bar and 6 bar. These data were obtained by carefully regulating the flow rate and the load of the system (i.e., the resistors). As seen in the figure, the expander efficiency increases with increasing superheating level, because the density and the collision frequencies among molecules of the working fluid is reasonably reduced as increasing the superheating. As a result, the entropy generation during the expansion process is simultaneously reduced. Similar tendency can be found in the publication of Bracco et al. [22]: as the superheating was operated between 10 and 25 °C and at fixed rotational speed of the pump, the expander efficiency, cycle efficiency and power output simultaneously increased, but this trend became insignificant as the superheating was higher than 40 °C in a fixed expansion ratio from their experiment. In this experiment, the maximum efficiency of the expander is 76.8% at a system pressure difference of 5 bars and 72.2% at 6 bars. Similarly, the ORC efficiency increases with the superheating of the expander. The maximum cycle efficiency
4.2. Effect of superheating at expander inlet on the system performance This section discusses the influence of superheating of the expander versus the overall performance of the system. The cycle was fed with hot water at a pressure of one atmosphere. The
Fig. 10. Effect of superheating at expander inlet on expander and cycle performance. The solid symbols represent expander efficiency, and hollow symbols represent the cycle efficiency.
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of 8.38% was obtained with superheating at 14.2 °C and a cycle pressure difference of 6 bar, whereas 8.12% was obtained with 19.7 °C and 5 bar. During the experimental phase, the efficiency of the pump did not obviously vary with the superheating because of the pump performance is mainly a function of the operating pressure difference. Two experimental results are compared in Fig. 11, which based on the fixed pressure difference of 6 bars, and operated at the superheating of 3.3 °C and 14.2 °C, respectively. This figure reveals that as the cycle is operated at higher superheating, the pinch point temperature difference of the evaporator is higher than that of lower superheating, whereas the deviation of the pinch point temperature difference of the condenser is very small as the cycle is operated at different superheating. One must be reminded that the system was imposed higher refrigerant flow rate as the cycle was operated at a lower superheating. The record of the mass flow rate of 0.075 kg/s and 0.0708 kg/s was obtained as the superheating was operated at 3.3 °C and 14.2 °C, respectively. Fig. 11 shows that as the superheating is operated at 3.3 °C, the temperature difference between hot water inlet and expander inlet is higher 1.2 °C than operated at 14.2 °C. As the cycle is operated at the superheating of 14.2 °C, the specific irreversibility of the evaporator is bigger than that of 3.3 °C, the ‘‘redundant specific irreversibility’’ is depicted in the gray area as the superheating were operated at 14.2 °C and 3.3 °C, respectively. In the other words, the specific irreversibility of the evaporator during the heat exchange process is relatively higher, if the cycle is operated at higher superheating. On the other hand, the difference of the specific irreversibility of the condenser is not significant as the cycle operated at the superheating between 14.2 °C and 3.3 °C.
4.3. The integration of experimental performance for different built-in volume ratio of scroll expanders The integrated experimental performance with respect to different scroll-type expanders are shown in Fig. 12(a), the
Fig. 12. Experimental performance based on different built-in volume ratio of the scroll expander (a) expander power output versus cycle efficiency; (b) heat source temperature versus operating pressure difference of the cycle.
Fig. 11. T-s diagram of the cycle operated with superheating at 3.3 °C (red) and 14.2 °C (blue). The red cycle represents the expander operated at 2384 RPM, and blue cycle represents the expander operated at 2433 RPM. (For interpretation of the references to color in this figure legend, the reader is referred to the web version of this article.)
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corresponding heat source temperature and operating pressure difference of the system in each experiment are shown in Fig. 12(b). In this figure, symbols Vr2 and Vr3 express the different built-in volume ratio of the scroll type expander obtained by previous test [19]. In previous experiment, the cycle was also fed with hot water at a pressure of one atmosphere as the heat transfer fluid. The previous experimental results involve rotational speed between 1641 RPM and 3054 RPM of the expander and a fixed superheating of 3 ± 1 K. Fig. 12 conveys that as the cycle is operated at the higher pressure difference, the system presents predominant performance as the scroll expander with bigger built-in volume ratio is utilized. In view of this, the expander used in the present study is inadequate to operate at a pressure difference of 5 bars in the system, because the output power of the expander is reduced to prohibitive value. Based on the integrated experimental data, the expander used in present study is preferred to operate at the pressure difference higher than 6 bars. 5. Conclusions This study experimentally investigates the performance of a scroll expander with a built-in volume ratio of 4.05 in an ORC system using HFC-245fa as working fluid, and the fluid has been mixed with moderate refrigerant oil that circulated in the system. The expander was modified from an oil-free open-drive scroll type air compressor and operated in reverse mode. The hot water at a pressure of one atmosphere was used as a dummy waste heat source and a cooling tower served as a heat sink. The main findings are as follows: 1. As the cycle is operated in fixed superheating of 3 ± 1 K, the maximum cycle efficiency is 9.43% at a rotational speed of 1535 ± 40 RPM and a cycle pressure difference of 8 bars. The maximum power of the expander is 2.3 kW at a rotational speed of 2970 ± 50 RPM and an operating cycle pressure difference of 8 bars. 2. The expander efficiency decreases as the rotational speed increases. A maximum expander isentropic efficiency is 73% at a rotational speed of 1535 ± 40 RPM and a cycle pressure difference of 8 bars. As the expander is operated between 1535 and 2130 RPM, the isentropic efficiency almost remains identical under above rotational speeds. 3. The maximum power produced by the expander is not corresponding to the maximum cycle efficiency for the real cycle. The compromise operating conditions have relation with system pressure difference and rotational speed of the expander. 4. Both efficiencies of the expander and the cycle increase with the increasing of superheating, but the irreversibility of the evaporator also increase simultaneously. 5. Experimental data from the scroll type expanders with different built-in volume ratio have been integrated. The expander used in this study is preferred to operate at higher range of pressure difference.
Acknowledgements The authors would like to appreciate the financial supports provided by Ministry of Science and Technology of the Republic of China, Taiwan (Contract No. MOST 103-2221-E-027-106); National Taipei University of Technology, Taipei, Taiwan (Contract No. 103T140); especially appreciate Kwan Chiu Radio Manufacturing Co., Ltd. for their experimental place and technical supports.
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