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Parallel-expander Organic Rankine cycle using dual expanders with different capacities
Energy 113 (2016) 204e214
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Parallel-expander Organic Rankine cycle using dual expanders with different capacities Eunkoo Yun a, b, Dokyun Kim b, Minseog Lee b, Seungdong Baek b, Sang Youl Yoon c, **, Kyung Chun Kim b, * a b c
Thermal Hydraulics Safety Research Division, Korea Atomic Energy Research Institute, 1045 Daedeokdaero, Yuseong, Deajeon 305-353, South Korea School of Mechanical Engineering, Pusan National University, Busan 609-735, South Korea Rolls-Royce and Pusan National University Technology Centre, Pusan National University, Busan 609-735, South Korea
a r t i c l e i n f o
a b s t r a c t
Article history: Received 2 March 2016 Received in revised form 7 June 2016 Accepted 7 July 2016
This study reports on a parallel-expander Organic Rankine cycle (PE-ORC) using dual expanders with different capacities. This system is applicable to waste heat sources that have large variation, such as distributed power plants with varying electricity demand. The test bench for the experimental investigation consists of a PE-ORC loop with two different scroll expanders, an electrical water heating system, and an air-cooled chiller. The tested performance characteristics for each operating mode are presented. The experimental results show three clearly separated operation regions with high efficiency, two possible switching points of the operating mode, and good heat recovery capability over a wider range of heat input compared to a PE-ORC system with two identical expanders. A control strategy for the system is proposed to maintain superheated state at the inlet of expander and to obtain high performance after changing the operating mode. © 2016 Elsevier Ltd. All rights reserved.
Keywords: Organic Rankine cycle Multiple parallel expanders Part-load Off-design Waste heat variation Internal combustion engine (ICE) Waste heat recovery (WHR) Scroll expander R245fa
1. Introduction Recently, the use of distributed power generation has been increasing to meet local power demand. Internal combustion engines (ICEs) are widely employed in distributed power plants due to their high electrical efficiency, reliability, and wide range of electrical power [1]. However, over 40% of the energy consumed in an ICE is still emitted to the atmosphere in the form of exhaust gas and jacket coolants [2,3]. Therefore, applying a waste heat recovery (WHR) system could enhance the overall efficiency of ICEs. The Organic Rankine cycle (ORC) could be considered for the heat recovery of an ICE system because it can act as a bottoming power cycle with a simple structure. A distributed power plant usually employs multiple ICE units for flexible operation to meet the varying electricity demand efficiently [4]. Since the flexible
* Corresponding author. ** Corresponding author. E-mail addresses: [email protected] (K.C. Kim).
(S.Y.
http://dx.doi.org/10.1016/j.energy.2016.07.045 0360-5442/© 2016 Elsevier Ltd. All rights reserved.
Yoon),
[email protected]
operation produces high variation of the total amount of waste heat, a WHR system for a distributed power plant should be designed with consideration of the large variation of the heat source and the overall efficiency of the power plant. Previous studies have focused on the WHR system for a single-ICE unit [5,6]. Thus, a WHR system must be developed for flexibly operated distributed power plants. Fig. 1 shows a conceptual illustration of a distributed power plant consisting of multiple ICE units with a WHR system based on an ORC system. The ICE units in the power plant are selectively operated by on/off control depending on the electricity demand, which is generally classified into base loads, intermediate loads, and peak loads. The amount of waste heat can be divided into roughly three classifications that depend on the electrical load, as shown in Fig. 1. These levels can be considered in the WHR system design for a distributed power plant. The ORC is a well-known and mature technology for obtaining useful power from low-temperature thermal energy, such as solar thermal energy [7e9], and geothermal energy [10,11], heat from biomass [12,13], and waste heat [5,6,14]. The ORC has also been
E. Yun et al. / Energy 113 (2016) 204e214
Nomenclature h
h
m_ N P Q_ T
t Vsw v w _ W
specific enthalpy, kJ/kg efficiency, % mass flow rate, kg/s rotational speed, rpm pressure, bar heat transfer rate, kW temperature, C torque, Nm swept volume, m3 specific volume, m3/kg specific work, kJ/kg power, kW
205
Subscripts C cooling water c cycle cd condenser ev evaporator ex expander H heat source i isentropic in inlet int internal out outlet pp pump r working fluid sh shaft sup superheating tot total
Fig. 1. Conceptual illustration of a PE-ORC for waste heat recovery in an ICE power plant.
considered as a possible candidate for waste heat recovery in distributed power plants. Previous studies on ORC for distributed power plants have focused on enhancing the ORC system performance with approaches such as working fluid selection [15,16], various cycle configurations [17e19], and using working fluid mixtures [20e22]. Most previous studies have been based on a basic ORC system with a single expander. However, Yun et al. [23] proposed the concept of a parallel-expander ORC (PE-ORC) to provide a design option for large variation in the amount of waste
heat produced. They showed the performance characteristics and the potential of the PE-ORC with two identical expanders with two design points. When two identical expanders are employed for a PE-ORC, a limited design for two heat source conditions corresponding to the maximum heat input and the half of them is possible. However, the distributed power plant shown in Fig. 1 has three representative waste heat conditions depending on the electrical load. Therefore, the PE-ORC applying two identical expanders is insufficient to meet
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such heat input variation. This study suggests a PE-ORC using dual expanders with different capacities, which is expected to provide three design points corresponding to base, intermediate, and peak loads. In addition, the system could have good heat recovery capability over a wider range of heat input than a system with two identical expanders or a basic single-expander ORC system. A control strategy for the PE-ORC is also proposed for keeping the state of expander inlet as superheated and providing high performance.
the specific volume and the pressure of the working fluid at the end of the isentropic expansion process. vint can be calculated by multiplying the built-in volume ratio of expander (rv) and the specific volume of expander inlet (vin,ex). In Eq. (2), the internal mass flow rate (m_ int ) can be determined as the volume flow rate divided by the specific volume of expander inlet.
m_ int ¼
Nex $Vsw;ex V_ ex ¼ vin;ex 60$vin;ex
(3)
2. Operating modes of PE-ORC The PE-ORC employs multiple expanders in parallel, whereas a basic ORC generally has a single expander. The PE-ORC system allows multiple operating modes, depending on the number of expanders operated. Yun et al. [23] experimentally investigated a simple PE-ORC system using two identical expanders in a previous study. They showed that the system can operate in “DUAL MODE” and “SINGLE MODE.” The PE-ORC can improve the heat recovery performance for large heat variation by changing the operating mode for a high pressure ratio. The two expanders with different capacities allows three operating modes to be realized in a single ORC loop, as shown in Fig. 2. When a large amount of heat is produced, the PE-ORC operates in “DUAL MODE” by opening both inlet valves. If the amount of heat is reduced, the PE-ORC can be selectively operated in “SINGLE MODE 1” or “SINGLE MODE 2” by open/close control of the inlet valves based on the amount of heat. The number of operating modes depends on the number of expanders and their capacities. Considering the two operating cases (on/off) for each expander and the non-operating case for the system, the maximum number of operating modes (Nop,max) can be simply determined as:
Nop;max ¼ 2n 1
(1)
where n is the number of expanders in the PE-ORC. The system performance for each operating mode could be obtained using the pre-determined performance characteristics of the expanders. Fig. 3 shows the simulated performance curves for a PE-ORC using an empirical model for the two scroll expanders [24]. _ The internal expansion power (W int;ex ) for each expander can be obtained by multiplying the internal mass flow rate (m_ int ) and the internal work (wint,ex).
_ _ int $wint;ex W int;ex ¼ m ¼ m_ int $ hin;ex hint þ vint $ Pint Pout;ex
(2)
where hint, vint and Pint denote, respectively, the specific enthalpy,
where Nex and Vsw,ex denote the rotational speed of expander and the swept volume of expander. The capacity of expander depends on the parameters in Eqs. (2) and (3). In this study, swept volume of expander is selected as a key parameter to determine the capacity of expander. Two PE-ORCs with expander capacity ratios of 1:1 and 2:1 are simulated and compared to each other. As presented in Fig. 3, the characteristics in “SINGLE MODE” have significant differences in the working range and the maximum capacity, while the performance characteristics for both cases in “DUAL MODE” are identical. 3. Experimental investigation of PE-ORC using scroll expanders 3.1. Experimental setup A PE-ORC system with two scroll expanders with different capacities has been designed for the experimental investigation. As shown in Fig. 4, the system mainly consists of two scroll expanders with different capacities, a working fluid feed pump, an evaporator, and a condenser. The specifications of the expanders are summarized in Table 1. Expander 1 is modified from an oil-free open drive air compressor by reversing the flow direction. A sealing container with a lip-type seal for Expander 1 was designed and manufactured to prevent leakage. A commercial 1-kW scroll expander manufactured by Air Squared, Inc., was used for Expander 2. Fig. 5 shows photographs of the expanders in the PE-ORC. Both scroll expanders are directly connected to a torque sensor and a permanent magnetic motor. The rotational speeds of the expanders were controlled by motor drivers. A volumetric pump was used for the working fluid, and platetype heat exchangers were used as the evaporator and the condenser. As a heat source, pressurized hot water produced by a 150-kW electric heater is supplied to the evaporator. Cooling water is used as a heat sink and fed from an air-cooled chiller into the condenser. The temperatures of both the heat source and heat sink can be controlled by the simple on/off controllers in each device.
Fig. 2. Three operating modes of PE-ORC using dual expanders with different capacities: (a) DUAL MODE, (b) SINGLE MODE 1, (c) SINGLE MODE 2. (a) PE-ORC with two identical expanders. (b) PE-ORC with two different expanders (capacity ratio ¼ 2:1).
E. Yun et al. / Energy 113 (2016) 204e214
207
1 DUAL MODE (ex1:on / ex2:on) SINGLE MODE (ex1:on / ex2:off)
Wex,total [-]
0.8 0.6 0.4 0.2 0
0
0.2
0.4
0.6
0.8
1
0.6
0.8
1
Q ev [-]
(a) PE-ORC with two identical expanders 1 DUAL MODE (ex1:on (EX1:ON/ ex2:on) / EX2:ON) SINGLE MODE1 (ex1:on / ex2:off) SINGLE MODE2 (ex1:off / ex2:on)
Wex,total [-]
0.8 0.6 0.4 0.2 0
0
0.2
0.4
Q ev [-]
(b) PE-ORC with two different expanders (capacity ratio = 2:1) Fig. 3. Simulated performance curves of PE-ORCs depending on expander capacity ratios.
CONTROLLER
EXPANDER 1 T T
EVAPORATOR
T
P
T T
F
T
FREQUENCY DRIVER
T
PM MOTOR
M
P
P
EXPANDER 2
T
M
T
P
C
T
CONDENSER
P
P
LIQUID RECEIVER
P
T
T
P
T
FREQUENCY DRIVER
PUMP
RESISTANCE
MOTOR DRIVER
P
T
F
PUMP
FREQUENCY DRIVER
P
PUMP P
ELECTRIC HEATER (Pressurized hot water) TEMPERATURE CONTROLLER
VALVE (OPEN/CLOSE) T
THERMOCOUPLE
P
PRESSURE TRANSMITTER
F
VOLUME FLOW METER
C
CORIOLIS MASS FLOW METER
M
TORQUE METER
AIR-COOLED CHILLER (Cooling water) TEMPERATURE CONTROLLER
Fig. 4. Test bench of PE-ORC (a) Expander 1 (b) Expander 2.
The flow rates of the heat source, heat sink, and working fluid are controlled by frequency drivers. The specifications of the sensors used in the test bench are listed in Table 2, and the corresponding
measurement points are presented in Fig. 4. Pressure transmitters and k-type thermocouples were positioned at the inlets and outlets of the main components, and a Coriolis flow meter was installed to
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E. Yun et al. / Energy 113 (2016) 204e214 Table 1 Specifications of expanders used in PE-ORC test bench.
Model numbers Manufacturer Max. pressure Expansion ratio Swept volume/rev
Expander 1 (modified from air compressor)
Expander 2
BC-KL52H Kyungwon Machinery Co, LTD. 10 bar (Compressed air) 4.05 25 cm3/rev
E15H22N4.25 Air Squared, Inc. 13.8 bar 3.5 12 cm3/rev
Fig. 5. Expanders in the PE-ORC test bench. (a) Expander isentropic efficiencies against pressure ratio over the expander. (b) Cycle efficiencies against pressure ratio over the expander.
Table 2 Specifications of sensors used in the test bench. Measurement
Type
Range
Accuracy
Pressure Mass flow rate Rotational speed Torque
Piezo resistive Coriolis flow meter Magneto type Strain gauge
0e5 bar/0e15 bar 0e0.305 kg/s 0e10000 rpm 0e50 Nm
±0.5% F.S. ±0.1% Reading ±1 rpm ±1% F.S.
measure the mass flow rate of the working fluid. In this study, R245fa was selected as the working fluid due to its suitable characteristics for ORC system in thermodynamic, environmental and safety criteria [15]. Since the condensing pressure of R245fa is similar to the atmospheric pressure at ambient temperature, the air infiltration problem into the cycle can be avoided. 3.2. Performance characteristics of operating modes The experimental conditions for characterization of the three operating modes are summarized in Table 3. The performance of an expander mainly depends on the rotational speed, the swept volume, and the built-in volume ratio. Thus, for all cases, the rotational speeds of Expander 1 and Expander 2 were set to 2000 rpm and 2200 rpm, respectively. This results in an expander capacity ratio of about 2:1. The expander inlet pressure ranged from 5 to 12 bar, which corresponds to pressure ratios of 3.5 and 6.5. The hot water and cooling water temperatures were fixed as 120 C and 25 C, Table 3 Experimental conditions of the performance tests for the three operating modes. Operating mode
Operation of expanders
SINGLE MODE 1
Expander Expander Expander Expander Expander Expander
SINGLE MODE 2 DUAL MODE
1 2 1 2 1 2
ON OFF OFF ON ON ON
RPM
TH,in ( C)
TC,in ( C)
Pex,in (bar)
2000 e e 2200 2000 2200
120
25
5e12
respectively. The shaft power is calculated by multiplying the measured rotational speed (Nex) of the expander by the measured shaft torque (t):
2pNex _ W ex;sh ¼ t$ 60
(6)
The total shaft power is determined by the sum of both shaft powers:
_ _ _ W ex;tot;sh ¼ W ex1;sh þ W ex2;sh
(5)
The maximum shaft powers measured in DUAL MODE, SINGLE MODE 1, and SINGLE MODE 2 were about 2.9 kW, 1.9 kW, and 1.1 kW, respectively. The isentropic efficiency of an expander (hi) and the cycle efficiency (hc) were calculated using Eqs. (7) and (8), respectively:
hi ¼
_ W ex;tot;sh m_ r $Dhi _ W
hc ¼ _ net ¼ Q ev
(7)
_ _ W ex;tot;sh W pp _ Q
(8)
ev
where m_ r is the mass flow rate of the working fluid, Dhi is the enthalpy difference between the expander inlet and outlet for _ pp is the power consumption through the isentropic expansion, W working fluid pump, and Q_ ev is the evaporative heat transfer rate. The evaporative heat transfer rate (Q_ ) was determined as: ev
Q_ ev ¼ m_ r hev;out hev;in
(9)
Due to the absence of power meter, the power consumption of _ pp ) was approximated by using a conthe working fluid pump (W stant isentropic efficiency (hpp) and measured properties at inlet and outlet of pump. Assuming the working fluid passing through _ pp can be calculated by: the pump is very low compressibility, W
E. Yun et al. / Energy 113 (2016) 204e214
_ pp W
. ¼ m_ r vpp;in Ppp;out Ppp;in hpp
(10)
where hpp is assumed to be a fixed value of 75% for all cases. The thermodynamic properties at measuring points were calculated in EES environmental [25]. Fig. 6 shows the performance test results for the three operating modes. The error bars represent measurement uncertainties obtained by uncertainty propagation theory [26]. The pressure drops in the evaporator and the condenser were less than 5 kPa for all cases. Therefore, the pressure drops caused by the heat exchangers and pipe system could be negligible in the present system. The condensing temperature was ranged from 25.5 to 27.8 C in both SINGLE MODEs. However, higher condensing temperature up to 31.5 C is observed in DUAL MODE. Those results clearly showed that the condensing pressure is proportional to the mass flow rate
209
of working fluid when cooling water temperature is constant. Fig. 6(a) shows the isentropic efficiency of an expander (hi) with respect to the pressure ratio for each operating mode. The maximum efficiency for SINGLE MODE 1 is about 63% at a pressure ratio of 5.8, and that for SINGLE MODE 2 is about 65% at a pressure ratio of 3.4. The cycle efficiencies for the three operating modes are shown in Fig. 6(b). A higher cycle efficiency is observed in SINGLE MODE 2 than in SINGLE MODE 1 when the pressure ratio is less than 4.7. The isentropic and cycle efficiencies of DUAL MODE are similar to the average efficiency between SINGLE MODE 1 and SINGLE MODE 2, as shown in the figure. The solid line superimposed on the test data is a fitted curve obtained with a third-order polynomial for each operation. The expected performance for DUAL MODE is represented with a dashed line that was calculated by averaging between the fitted curves for SINGLE MODE 1 and SINGLE MODE 2. The test results show that the measured isentropic and cycle efficiencies for DUAL
Fig. 6. Performance characteristics for three operating modes.
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MODE closely matched with the expected curves. Thus, the performance characteristics in this mode can be predicted using the performance characteristics of the selected expanders. In Fig. 7, the test results are replotted against the evaporative heat transfer rate over the whole range to identify the characteristics of the PE-ORC more clearly in regard to heat recovery capability [23]. The figure shows the three clearly separated operation regions with high efficiency in heat recovery and two possible switching points of the operating mode. The result reveals that the proposed PE-ORC system could have good heat recovery capability over a wider range of heat input than a system with two identical expanders or a basic single-expander system. The operating mode could be switched at evaporative heat transfer rates of about 11 kW and 20 kW. 4. Control strategy during the change of operating mode 4.1. Control strategy The general control strategy of a basic ORC in part-load conditions is to maintain constant superheating temperature at the entrance of the expander. For the control of the PE-ORC system, the
changing direction of the operating mode from high to low capacity and from low to high capacity should be considered. When the operating mode of a PE-ORC changes from high capacity to low capacity with low superheating (i.e., DUAL MODE to SINGLE MODE 1 or from SINGLE MODE 1 to SINGLE MODE 2), the pressure at the expander inlet could increase excessively over the evaporating pressure. Liquid droplets could be generated by the excessive pressure increase before the entrance of an expander and significantly damage the expanders. Thus, a control strategy for sufficient superheating at the entrance of the expander during operating mode change should be investigated. For sufficient superheating, one control method is to reduce the flow rate of the working fluid to decrease the evaporating pressure before the operating mode change. The control method was conceptually simulated for a case of operating mode change from DUAL MODE to SINGLE MODE 1, and its effects are shown in Fig. 8. To predict the evaporating pressure and superheating temperature variations caused by pump and operating mode controls, a heat exchanger model with generalized moving-boundary algorithm [27] is combined with the scroll expander model. The simulation was performed in EES environmental. In this simulation, the operational superheating temperature and the heat source
3.5 3
Wex [kW]
2.5
SINGLE MODE1 (measured) (ex1=ON / ex2=OFF) SINGLE MODE2 (ex1=OFF (measured)/ ex2=ON) DUAL MODE (measured) DUAL MODE (ex1=ON / ex2=ON) Fitted curve from SINGLE MODE1 Fitted curve from SINGLE MODE2 Fitted curve from DUAL MODE
2 1.5 1 0.5 0 9
ηc [%]
8 7 6 5
SINGLE MODE1 (measured) (ex1=ON / ex2=OFF) SINGLE MODE2 (measured) (ex1=OFF / ex2=ON) DUAL MODE (measured) DUAL curve MODEfrom (ex1=ON / ex2=ON) Fitted SINGLE MODE1 Fitted curve from SINGLE MODE2 Fitted curve from DUAL MODE
4 3 0
5
10
15
20
25
30
35
40
Qev [kW] Fig. 7. Performance characteristics against evaporative heat transfer (a) without control, (b) with control.
E. Yun et al. / Energy 113 (2016) 204e214
temperature were assumed to 10 C and 120 C, respectively. Fig. 8(a) shows that after the mode change without any pump control, an expander inlet vapor quality less than 1.0 could result. Thus, the flow rate of the working fluid should be reduced before the mode change. The control process is presented in a T-S diagram in Fig. 8(b). Before the mode change, reducing the flow rate decreases the expander inlet pressure and increases the superheating (process 1e20 in Fig. 8(b)). After the mode change, sufficient superheating could be maintained even though the expander inlet pressure is increased (process 20 e30 in Fig. 8(b)). When the operating mode of the PE-ORC system changes from low capacity to high capacity, the pressure at the expander inlet decreases, and the superheating could be larger than that before
211
the mode change (process 30 e20 in Fig. 8(b)). Thus, the vapor quality would not be an issue, and the sufficient superheating allows the expander inlet pressure to be increase for system performance (process 20 e1 in Fig. 8(b)). The control method in this case is to increase the expander inlet pressure by increasing the flow rate of working fluid after the mode change. 4.2. Test for operating mode changes Yun et al. [23] investigated the general behaviors of the total power output during operating mode change for PE-ORC using two identical expanders. The degree of superheat before the expander inlet is an important control parameter for preventing expander
Fig. 8. T-s diagram for operating mode change. (a) Total shaft power. (b) Cycle efficiency.
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E. Yun et al. / Energy 113 (2016) 204e214
damage, so this study focuses on the superheating variation when changing from high capacity mode to low capacity mode (from DUAL MODE to SINGLE MODE 1 and from SINGLE MODE 1 to SINGLE MODE 2). The rotational speeds of both expanders were made the same as in the performance characterization test. All tests were carried out with a fixed rotational speed of the working fluid pump for each case. A degree of superheat before operating mode changes was considered to be higher than 25 C based on the vapor quality drop after the mode change. The test results in Fig. 9 are superimposed on the performance curves from Fig. 6 and are summarized in Table 4. The results reveal good reproducibility of the performance curves. For both cases, the total shaft power and the cycle efficiency increased after each mode change, while the evaporative heat transfer rate decreased. The
Table 4 Data recorded during the dynamic tests. Case
Operating mode change
1
DUAL MODE to SINGLE MODE 1 SINGLE MODE 1 to SINGLE MODE 2
2
Q_ ev
_ ex;tot W
hc
(kW)
(kW)
(%)
(bar)
(kg/s)
(K)
20.84 19.22 10.44 9.406
1.21 1.56 0.52 0.73
5.88 7.92 4.76 7.35
6.75 10.65 5.96 8.99
0.07802 0.07382 0.03903 0.03591
26.38 11.43 28.32 11.96
Pex,in
m_ r
DTsup
evaporative heat transfer rate was slightly reduced by the reduced flow rate of working fluid due to the rapid increase of the pump outlet pressure. The meaningful increases in total shaft power are produced by the increased pressure ratio.
Fig. 9. Mode change test results superimposed on performance curves from Fig. 7. (a) Case 1 (b) Case 2.
E. Yun et al. / Energy 113 (2016) 204e214
Fig. 10. Temporal variations in pressures over the expanders, mass flow rate of working fluid, superheating at expander inlet and evaporative heat transfer.
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Fig. 10 presents the temporal variations in inlet/outlet pressures, mass flow rate, degree of superheat and evaporative heat transfer rate during mode change tests. For both cases, there is a significant drop of more than 15 C in the superheating as the operating mode changes from high to low capacity. Thus, the mass flow rate of working fluid should be reduced appropriately for sufficient superheating before the operating mode change to maintain the state of expander inlet as superheated after the change. After the mode change, the mass flow rate of the working fluid decreases slightly due to the significant increase of the pump outlet pressure. This decrease depends on the pump performance and could be considered when determining the flow rate reduction for the operating mode change. While the enthalpy difference through the evaporator is almost constant during the mode changes, the evaporative heat transfer rate is slightly reduced after the operating mode changes. This depicts that the variation in evaporative heat transfer is strongly affected by the variation in mass flow rate of working fluid during the operating mode changes.
5. Conclusions This study experimentally investigated the performance characteristics and control methods for a PE-ORC system using two expanders with different capacities. A test bench was built with a PE-ORC loop and two different scroll expanders, an electrical water heating system, and an air-cooled chiller. The maximum number of operating modes (Nop,max) is 2n1 for n expanders in the PE-ORC. The number of operating modes was determined to be 3 due to the different expander capacities for DUAL MODE, SINGLE MODE 1, and SINGLE MODE 2. The results of the characterization test showed that the performance in DUAL MODE closely matched the average expander performance and could be used to design a PE-ORC system. The tested performance curves against the evaporative heat transfer rate showed three clearly separated operating regions with high efficiency and two possible switching points for the operating mode. The test results reveal that the proposed system could have a good heat recovery capability over a wide variation of heat input. Control strategies for the PE-ORC system were analytically and experimentally studied. The results showed that a significant drop in the superheating occurred after the operating mode changed from high to low capacity. The expander inlet pressure should thus be reduced appropriately for sufficient superheating before the operating mode change by decreasing the flow rate of working fluid. Control methods were proposed for PE-ORC operation to maintain the state of expander inlet as superheated and achieve high-performance heat recovery in PE-ORC system.
Acknowledgements This research was partially supported by Basic Science Research Program through the National Research Foundation of Korea (NRF) funded by the Ministry of Education (NRF-2013R1A1A2012173), and partially by the National Research Foundation of Korea (NRF) grant funded by the Korea Government (MSIP) through GCRC-SOP (No. 2011-0030013).
References [1] Vaja I, Gambarotta A. Internal combustion engine (ICE) bottoming with organic rankine cycles (ORCs). Energy 2010;35:1084e93. [2] Tchanche BF, Lambrinos G, Frangoudakis A. Low-grade heat conversion into power using organic Rankine cyclesea review of various applications. Renew Sustain Energy Rev 2011;15:3963e79. [3] Singh DV, Pedersen E. A review of waste heat recovery technologies for maritime applications. Energy Convers Manag 3/1/2016;111:315e28. € der L. Distributed generation: a definition1. [4] Ackermann T, Andersson G, So Electr Power Syst Res 4/20/2001;57:195e204. [5] Choi BC, Kim YM. Thermodynamic analysis of a dual loop heat recovery system with trilateral cycle applied to exhaust gases of internal combustion engine for propulsion of the 6800 TEU container ship. Energy 2013;58: 404e16. [6] Yang F, Zhang H, Song S, Bei C, Wang H, Wang E. Thermoeconomic multiobjective optimization of an organic Rankine cycle for exhaust waste heat recovery of a diesel engine. Energy 12/15/2015;93(Part 2):2208e28. [7] Kim K, Han C. A review on solar collector and solar organic rankine cycle (ORC) systems. J Autom Control Eng 2015;3(1):66e73. [8] Wang JL, Zhao L, Wang XD. An experimental study on the recuperative low temperature solar Rankine cycle using R245fa. Appl Energy 2012;94:34e40. [9] Quoilin S, Orosz M, Hemond H, Lemort V. Performance and design optimization of a low-cost solar Organic Rankine cycle for remote power generation. Sol Energy 2011;85(5):955e66. [10] Baik Y-J, Kim M, Chang K-C, Lee Y-S, Yoon H-K. Power enhancement potential of a mixture transcritical cycle for a low-temperature geothermal power generation. Energy 2012;47:70e6. [11] Shengjun Z, Huaixin W, Tao G. Performance comparison and parametric optimization of subcritical Organic Rankine cycle (ORC) and transcritical power cycle system for low-temperature geothermal power generation. Appl Energy 2011;88:2740e54. [12] Huang Y, McIlveen-Wright DR, Rezvani S, Huang MJ, Wang YD, Roskilly AP, et al. Comparative techno-economic analysis of biomass fuelled combined heat and power for commercial buildings. Appl Energy 2013;112:518e25. [13] Uris M, Linares JI, Arenas E. Size optimization of a biomass-fired cogeneration plant CHP/CCHP (Combined heat and power/Combined heat, cooling and power) based on Organic Rankine cycle for a district network in Spain. Energy 2015;88:935e45. [14] Oluleye G, Jobson M, Smith R, Perry SJ. Evaluating the potential of process sites for waste heat recovery. Appl Energy 1/1/2016;161:627e46. [15] Amicabile S, Lee JI, Kum D. A comprehensive design methodology of organic Rankine cycles for the waste heat recovery of automotive heavy-duty diesel engines. Appl Therm Eng 2015;87:574e85. [16] Tchanche BF, Papadakis G, Lambrinos G. Fluid selection for a low-temperature solar organic Rankine cycle. Appl Therm Eng 2009;29:2468e76. [17] Yamada N, Minami T, Mohamad MNA. Fundamental experiment of pumpless Rankine-type cycle for low-temperature heat recovery. Energy 2011;36(2): 1010e7. [18] Li X, Zhao C, Hu X. Thermodynamic analysis of organic rankine cycle with ejector. Energy 2012;42:342e9. [19] Song J, Gu C-w. Performance analysis of a dual-loop Organic Rankine cycle (ORC) system with wet steam expansion for engine waste heat recovery. Appl Energy 10/15/2015;156:280e9. [20] Bamorovat Abadi G, Yun E, Kim KC. Experimental study of a 1 kw Organic Rankine cycle with a zeotropic mixture of R245fa/R134a. Energy 12/15/ 2015;93(Part 2):2363e73. [21] Habka M, Ajib S. Evaluation of mixtures performances in Organic Rankine Cycle when utilizing the geothermal water with and without cogeneration. Appl Energy 9/15/2015;154:567e76. [22] Wang JL, Zhao L, Wang XD. A comparative study of pure and zeotropic mixtures in low-temperature solar Rankine cycle. Appl Energy 2010;87:3366e73. [23] Eunkoo Y, Dokyun K, Sang Youl Y, Kyung Chun K. Experimental investigation of an Organic Rankine cycle with multiple expanders used in parallel. Appl Energy 2015;145:246e54. [24] Quoilin S, Lemort V, Lebrun J. Experimental study and modeling of an Organic Rankine cycle using scroll expander. Appl Energy 2010;87(4):1260e8. [25] Klein SA, Alvarado FL. EES manual. Wisconsin, USA: Middleton; 1999. [26] Taylor BN, Kuyatt CE. Guidelines for evaluating and expressing the uncertainty of NIST measurement results. 1994 ed. Gaithersburg, MD: U.S. Department of Commerce, Technology Administration, National Institute of Standards and Technology; 1994. [27] Bell IH, Quoilin S, Georges E, Braun JE, Groll EA, Horton WT, et al. A generalized moving-boundary algorithm to predict the heat transfer rate of counterflow heat exchangers for any phase configuration. Appl Therm Eng 2015;79:192e201.
Contents lists available at ScienceDirect
Energy journal homepage: www.elsevier.com/locate/energy
Parallel-expander Organic Rankine cycle using dual expanders with different capacities Eunkoo Yun a, b, Dokyun Kim b, Minseog Lee b, Seungdong Baek b, Sang Youl Yoon c, **, Kyung Chun Kim b, * a b c
Thermal Hydraulics Safety Research Division, Korea Atomic Energy Research Institute, 1045 Daedeokdaero, Yuseong, Deajeon 305-353, South Korea School of Mechanical Engineering, Pusan National University, Busan 609-735, South Korea Rolls-Royce and Pusan National University Technology Centre, Pusan National University, Busan 609-735, South Korea
a r t i c l e i n f o
a b s t r a c t
Article history: Received 2 March 2016 Received in revised form 7 June 2016 Accepted 7 July 2016
This study reports on a parallel-expander Organic Rankine cycle (PE-ORC) using dual expanders with different capacities. This system is applicable to waste heat sources that have large variation, such as distributed power plants with varying electricity demand. The test bench for the experimental investigation consists of a PE-ORC loop with two different scroll expanders, an electrical water heating system, and an air-cooled chiller. The tested performance characteristics for each operating mode are presented. The experimental results show three clearly separated operation regions with high efficiency, two possible switching points of the operating mode, and good heat recovery capability over a wider range of heat input compared to a PE-ORC system with two identical expanders. A control strategy for the system is proposed to maintain superheated state at the inlet of expander and to obtain high performance after changing the operating mode. © 2016 Elsevier Ltd. All rights reserved.
Keywords: Organic Rankine cycle Multiple parallel expanders Part-load Off-design Waste heat variation Internal combustion engine (ICE) Waste heat recovery (WHR) Scroll expander R245fa
1. Introduction Recently, the use of distributed power generation has been increasing to meet local power demand. Internal combustion engines (ICEs) are widely employed in distributed power plants due to their high electrical efficiency, reliability, and wide range of electrical power [1]. However, over 40% of the energy consumed in an ICE is still emitted to the atmosphere in the form of exhaust gas and jacket coolants [2,3]. Therefore, applying a waste heat recovery (WHR) system could enhance the overall efficiency of ICEs. The Organic Rankine cycle (ORC) could be considered for the heat recovery of an ICE system because it can act as a bottoming power cycle with a simple structure. A distributed power plant usually employs multiple ICE units for flexible operation to meet the varying electricity demand efficiently [4]. Since the flexible
* Corresponding author. ** Corresponding author. E-mail addresses: [email protected] (K.C. Kim).
(S.Y.
http://dx.doi.org/10.1016/j.energy.2016.07.045 0360-5442/© 2016 Elsevier Ltd. All rights reserved.
Yoon),
[email protected]
operation produces high variation of the total amount of waste heat, a WHR system for a distributed power plant should be designed with consideration of the large variation of the heat source and the overall efficiency of the power plant. Previous studies have focused on the WHR system for a single-ICE unit [5,6]. Thus, a WHR system must be developed for flexibly operated distributed power plants. Fig. 1 shows a conceptual illustration of a distributed power plant consisting of multiple ICE units with a WHR system based on an ORC system. The ICE units in the power plant are selectively operated by on/off control depending on the electricity demand, which is generally classified into base loads, intermediate loads, and peak loads. The amount of waste heat can be divided into roughly three classifications that depend on the electrical load, as shown in Fig. 1. These levels can be considered in the WHR system design for a distributed power plant. The ORC is a well-known and mature technology for obtaining useful power from low-temperature thermal energy, such as solar thermal energy [7e9], and geothermal energy [10,11], heat from biomass [12,13], and waste heat [5,6,14]. The ORC has also been
E. Yun et al. / Energy 113 (2016) 204e214
Nomenclature h
h
m_ N P Q_ T
t Vsw v w _ W
specific enthalpy, kJ/kg efficiency, % mass flow rate, kg/s rotational speed, rpm pressure, bar heat transfer rate, kW temperature, C torque, Nm swept volume, m3 specific volume, m3/kg specific work, kJ/kg power, kW
205
Subscripts C cooling water c cycle cd condenser ev evaporator ex expander H heat source i isentropic in inlet int internal out outlet pp pump r working fluid sh shaft sup superheating tot total
Fig. 1. Conceptual illustration of a PE-ORC for waste heat recovery in an ICE power plant.
considered as a possible candidate for waste heat recovery in distributed power plants. Previous studies on ORC for distributed power plants have focused on enhancing the ORC system performance with approaches such as working fluid selection [15,16], various cycle configurations [17e19], and using working fluid mixtures [20e22]. Most previous studies have been based on a basic ORC system with a single expander. However, Yun et al. [23] proposed the concept of a parallel-expander ORC (PE-ORC) to provide a design option for large variation in the amount of waste
heat produced. They showed the performance characteristics and the potential of the PE-ORC with two identical expanders with two design points. When two identical expanders are employed for a PE-ORC, a limited design for two heat source conditions corresponding to the maximum heat input and the half of them is possible. However, the distributed power plant shown in Fig. 1 has three representative waste heat conditions depending on the electrical load. Therefore, the PE-ORC applying two identical expanders is insufficient to meet
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such heat input variation. This study suggests a PE-ORC using dual expanders with different capacities, which is expected to provide three design points corresponding to base, intermediate, and peak loads. In addition, the system could have good heat recovery capability over a wider range of heat input than a system with two identical expanders or a basic single-expander ORC system. A control strategy for the PE-ORC is also proposed for keeping the state of expander inlet as superheated and providing high performance.
the specific volume and the pressure of the working fluid at the end of the isentropic expansion process. vint can be calculated by multiplying the built-in volume ratio of expander (rv) and the specific volume of expander inlet (vin,ex). In Eq. (2), the internal mass flow rate (m_ int ) can be determined as the volume flow rate divided by the specific volume of expander inlet.
m_ int ¼
Nex $Vsw;ex V_ ex ¼ vin;ex 60$vin;ex
(3)
2. Operating modes of PE-ORC The PE-ORC employs multiple expanders in parallel, whereas a basic ORC generally has a single expander. The PE-ORC system allows multiple operating modes, depending on the number of expanders operated. Yun et al. [23] experimentally investigated a simple PE-ORC system using two identical expanders in a previous study. They showed that the system can operate in “DUAL MODE” and “SINGLE MODE.” The PE-ORC can improve the heat recovery performance for large heat variation by changing the operating mode for a high pressure ratio. The two expanders with different capacities allows three operating modes to be realized in a single ORC loop, as shown in Fig. 2. When a large amount of heat is produced, the PE-ORC operates in “DUAL MODE” by opening both inlet valves. If the amount of heat is reduced, the PE-ORC can be selectively operated in “SINGLE MODE 1” or “SINGLE MODE 2” by open/close control of the inlet valves based on the amount of heat. The number of operating modes depends on the number of expanders and their capacities. Considering the two operating cases (on/off) for each expander and the non-operating case for the system, the maximum number of operating modes (Nop,max) can be simply determined as:
Nop;max ¼ 2n 1
(1)
where n is the number of expanders in the PE-ORC. The system performance for each operating mode could be obtained using the pre-determined performance characteristics of the expanders. Fig. 3 shows the simulated performance curves for a PE-ORC using an empirical model for the two scroll expanders [24]. _ The internal expansion power (W int;ex ) for each expander can be obtained by multiplying the internal mass flow rate (m_ int ) and the internal work (wint,ex).
_ _ int $wint;ex W int;ex ¼ m ¼ m_ int $ hin;ex hint þ vint $ Pint Pout;ex
(2)
where hint, vint and Pint denote, respectively, the specific enthalpy,
where Nex and Vsw,ex denote the rotational speed of expander and the swept volume of expander. The capacity of expander depends on the parameters in Eqs. (2) and (3). In this study, swept volume of expander is selected as a key parameter to determine the capacity of expander. Two PE-ORCs with expander capacity ratios of 1:1 and 2:1 are simulated and compared to each other. As presented in Fig. 3, the characteristics in “SINGLE MODE” have significant differences in the working range and the maximum capacity, while the performance characteristics for both cases in “DUAL MODE” are identical. 3. Experimental investigation of PE-ORC using scroll expanders 3.1. Experimental setup A PE-ORC system with two scroll expanders with different capacities has been designed for the experimental investigation. As shown in Fig. 4, the system mainly consists of two scroll expanders with different capacities, a working fluid feed pump, an evaporator, and a condenser. The specifications of the expanders are summarized in Table 1. Expander 1 is modified from an oil-free open drive air compressor by reversing the flow direction. A sealing container with a lip-type seal for Expander 1 was designed and manufactured to prevent leakage. A commercial 1-kW scroll expander manufactured by Air Squared, Inc., was used for Expander 2. Fig. 5 shows photographs of the expanders in the PE-ORC. Both scroll expanders are directly connected to a torque sensor and a permanent magnetic motor. The rotational speeds of the expanders were controlled by motor drivers. A volumetric pump was used for the working fluid, and platetype heat exchangers were used as the evaporator and the condenser. As a heat source, pressurized hot water produced by a 150-kW electric heater is supplied to the evaporator. Cooling water is used as a heat sink and fed from an air-cooled chiller into the condenser. The temperatures of both the heat source and heat sink can be controlled by the simple on/off controllers in each device.
Fig. 2. Three operating modes of PE-ORC using dual expanders with different capacities: (a) DUAL MODE, (b) SINGLE MODE 1, (c) SINGLE MODE 2. (a) PE-ORC with two identical expanders. (b) PE-ORC with two different expanders (capacity ratio ¼ 2:1).
E. Yun et al. / Energy 113 (2016) 204e214
207
1 DUAL MODE (ex1:on / ex2:on) SINGLE MODE (ex1:on / ex2:off)
Wex,total [-]
0.8 0.6 0.4 0.2 0
0
0.2
0.4
0.6
0.8
1
0.6
0.8
1
Q ev [-]
(a) PE-ORC with two identical expanders 1 DUAL MODE (ex1:on (EX1:ON/ ex2:on) / EX2:ON) SINGLE MODE1 (ex1:on / ex2:off) SINGLE MODE2 (ex1:off / ex2:on)
Wex,total [-]
0.8 0.6 0.4 0.2 0
0
0.2
0.4
Q ev [-]
(b) PE-ORC with two different expanders (capacity ratio = 2:1) Fig. 3. Simulated performance curves of PE-ORCs depending on expander capacity ratios.
CONTROLLER
EXPANDER 1 T T
EVAPORATOR
T
P
T T
F
T
FREQUENCY DRIVER
T
PM MOTOR
M
P
P
EXPANDER 2
T
M
T
P
C
T
CONDENSER
P
P
LIQUID RECEIVER
P
T
T
P
T
FREQUENCY DRIVER
PUMP
RESISTANCE
MOTOR DRIVER
P
T
F
PUMP
FREQUENCY DRIVER
P
PUMP P
ELECTRIC HEATER (Pressurized hot water) TEMPERATURE CONTROLLER
VALVE (OPEN/CLOSE) T
THERMOCOUPLE
P
PRESSURE TRANSMITTER
F
VOLUME FLOW METER
C
CORIOLIS MASS FLOW METER
M
TORQUE METER
AIR-COOLED CHILLER (Cooling water) TEMPERATURE CONTROLLER
Fig. 4. Test bench of PE-ORC (a) Expander 1 (b) Expander 2.
The flow rates of the heat source, heat sink, and working fluid are controlled by frequency drivers. The specifications of the sensors used in the test bench are listed in Table 2, and the corresponding
measurement points are presented in Fig. 4. Pressure transmitters and k-type thermocouples were positioned at the inlets and outlets of the main components, and a Coriolis flow meter was installed to
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E. Yun et al. / Energy 113 (2016) 204e214 Table 1 Specifications of expanders used in PE-ORC test bench.
Model numbers Manufacturer Max. pressure Expansion ratio Swept volume/rev
Expander 1 (modified from air compressor)
Expander 2
BC-KL52H Kyungwon Machinery Co, LTD. 10 bar (Compressed air) 4.05 25 cm3/rev
E15H22N4.25 Air Squared, Inc. 13.8 bar 3.5 12 cm3/rev
Fig. 5. Expanders in the PE-ORC test bench. (a) Expander isentropic efficiencies against pressure ratio over the expander. (b) Cycle efficiencies against pressure ratio over the expander.
Table 2 Specifications of sensors used in the test bench. Measurement
Type
Range
Accuracy
Pressure Mass flow rate Rotational speed Torque
Piezo resistive Coriolis flow meter Magneto type Strain gauge
0e5 bar/0e15 bar 0e0.305 kg/s 0e10000 rpm 0e50 Nm
±0.5% F.S. ±0.1% Reading ±1 rpm ±1% F.S.
measure the mass flow rate of the working fluid. In this study, R245fa was selected as the working fluid due to its suitable characteristics for ORC system in thermodynamic, environmental and safety criteria [15]. Since the condensing pressure of R245fa is similar to the atmospheric pressure at ambient temperature, the air infiltration problem into the cycle can be avoided. 3.2. Performance characteristics of operating modes The experimental conditions for characterization of the three operating modes are summarized in Table 3. The performance of an expander mainly depends on the rotational speed, the swept volume, and the built-in volume ratio. Thus, for all cases, the rotational speeds of Expander 1 and Expander 2 were set to 2000 rpm and 2200 rpm, respectively. This results in an expander capacity ratio of about 2:1. The expander inlet pressure ranged from 5 to 12 bar, which corresponds to pressure ratios of 3.5 and 6.5. The hot water and cooling water temperatures were fixed as 120 C and 25 C, Table 3 Experimental conditions of the performance tests for the three operating modes. Operating mode
Operation of expanders
SINGLE MODE 1
Expander Expander Expander Expander Expander Expander
SINGLE MODE 2 DUAL MODE
1 2 1 2 1 2
ON OFF OFF ON ON ON
RPM
TH,in ( C)
TC,in ( C)
Pex,in (bar)
2000 e e 2200 2000 2200
120
25
5e12
respectively. The shaft power is calculated by multiplying the measured rotational speed (Nex) of the expander by the measured shaft torque (t):
2pNex _ W ex;sh ¼ t$ 60
(6)
The total shaft power is determined by the sum of both shaft powers:
_ _ _ W ex;tot;sh ¼ W ex1;sh þ W ex2;sh
(5)
The maximum shaft powers measured in DUAL MODE, SINGLE MODE 1, and SINGLE MODE 2 were about 2.9 kW, 1.9 kW, and 1.1 kW, respectively. The isentropic efficiency of an expander (hi) and the cycle efficiency (hc) were calculated using Eqs. (7) and (8), respectively:
hi ¼
_ W ex;tot;sh m_ r $Dhi _ W
hc ¼ _ net ¼ Q ev
(7)
_ _ W ex;tot;sh W pp _ Q
(8)
ev
where m_ r is the mass flow rate of the working fluid, Dhi is the enthalpy difference between the expander inlet and outlet for _ pp is the power consumption through the isentropic expansion, W working fluid pump, and Q_ ev is the evaporative heat transfer rate. The evaporative heat transfer rate (Q_ ) was determined as: ev
Q_ ev ¼ m_ r hev;out hev;in
(9)
Due to the absence of power meter, the power consumption of _ pp ) was approximated by using a conthe working fluid pump (W stant isentropic efficiency (hpp) and measured properties at inlet and outlet of pump. Assuming the working fluid passing through _ pp can be calculated by: the pump is very low compressibility, W
E. Yun et al. / Energy 113 (2016) 204e214
_ pp W
. ¼ m_ r vpp;in Ppp;out Ppp;in hpp
(10)
where hpp is assumed to be a fixed value of 75% for all cases. The thermodynamic properties at measuring points were calculated in EES environmental [25]. Fig. 6 shows the performance test results for the three operating modes. The error bars represent measurement uncertainties obtained by uncertainty propagation theory [26]. The pressure drops in the evaporator and the condenser were less than 5 kPa for all cases. Therefore, the pressure drops caused by the heat exchangers and pipe system could be negligible in the present system. The condensing temperature was ranged from 25.5 to 27.8 C in both SINGLE MODEs. However, higher condensing temperature up to 31.5 C is observed in DUAL MODE. Those results clearly showed that the condensing pressure is proportional to the mass flow rate
209
of working fluid when cooling water temperature is constant. Fig. 6(a) shows the isentropic efficiency of an expander (hi) with respect to the pressure ratio for each operating mode. The maximum efficiency for SINGLE MODE 1 is about 63% at a pressure ratio of 5.8, and that for SINGLE MODE 2 is about 65% at a pressure ratio of 3.4. The cycle efficiencies for the three operating modes are shown in Fig. 6(b). A higher cycle efficiency is observed in SINGLE MODE 2 than in SINGLE MODE 1 when the pressure ratio is less than 4.7. The isentropic and cycle efficiencies of DUAL MODE are similar to the average efficiency between SINGLE MODE 1 and SINGLE MODE 2, as shown in the figure. The solid line superimposed on the test data is a fitted curve obtained with a third-order polynomial for each operation. The expected performance for DUAL MODE is represented with a dashed line that was calculated by averaging between the fitted curves for SINGLE MODE 1 and SINGLE MODE 2. The test results show that the measured isentropic and cycle efficiencies for DUAL
Fig. 6. Performance characteristics for three operating modes.
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E. Yun et al. / Energy 113 (2016) 204e214
MODE closely matched with the expected curves. Thus, the performance characteristics in this mode can be predicted using the performance characteristics of the selected expanders. In Fig. 7, the test results are replotted against the evaporative heat transfer rate over the whole range to identify the characteristics of the PE-ORC more clearly in regard to heat recovery capability [23]. The figure shows the three clearly separated operation regions with high efficiency in heat recovery and two possible switching points of the operating mode. The result reveals that the proposed PE-ORC system could have good heat recovery capability over a wider range of heat input than a system with two identical expanders or a basic single-expander system. The operating mode could be switched at evaporative heat transfer rates of about 11 kW and 20 kW. 4. Control strategy during the change of operating mode 4.1. Control strategy The general control strategy of a basic ORC in part-load conditions is to maintain constant superheating temperature at the entrance of the expander. For the control of the PE-ORC system, the
changing direction of the operating mode from high to low capacity and from low to high capacity should be considered. When the operating mode of a PE-ORC changes from high capacity to low capacity with low superheating (i.e., DUAL MODE to SINGLE MODE 1 or from SINGLE MODE 1 to SINGLE MODE 2), the pressure at the expander inlet could increase excessively over the evaporating pressure. Liquid droplets could be generated by the excessive pressure increase before the entrance of an expander and significantly damage the expanders. Thus, a control strategy for sufficient superheating at the entrance of the expander during operating mode change should be investigated. For sufficient superheating, one control method is to reduce the flow rate of the working fluid to decrease the evaporating pressure before the operating mode change. The control method was conceptually simulated for a case of operating mode change from DUAL MODE to SINGLE MODE 1, and its effects are shown in Fig. 8. To predict the evaporating pressure and superheating temperature variations caused by pump and operating mode controls, a heat exchanger model with generalized moving-boundary algorithm [27] is combined with the scroll expander model. The simulation was performed in EES environmental. In this simulation, the operational superheating temperature and the heat source
3.5 3
Wex [kW]
2.5
SINGLE MODE1 (measured) (ex1=ON / ex2=OFF) SINGLE MODE2 (ex1=OFF (measured)/ ex2=ON) DUAL MODE (measured) DUAL MODE (ex1=ON / ex2=ON) Fitted curve from SINGLE MODE1 Fitted curve from SINGLE MODE2 Fitted curve from DUAL MODE
2 1.5 1 0.5 0 9
ηc [%]
8 7 6 5
SINGLE MODE1 (measured) (ex1=ON / ex2=OFF) SINGLE MODE2 (measured) (ex1=OFF / ex2=ON) DUAL MODE (measured) DUAL curve MODEfrom (ex1=ON / ex2=ON) Fitted SINGLE MODE1 Fitted curve from SINGLE MODE2 Fitted curve from DUAL MODE
4 3 0
5
10
15
20
25
30
35
40
Qev [kW] Fig. 7. Performance characteristics against evaporative heat transfer (a) without control, (b) with control.
E. Yun et al. / Energy 113 (2016) 204e214
temperature were assumed to 10 C and 120 C, respectively. Fig. 8(a) shows that after the mode change without any pump control, an expander inlet vapor quality less than 1.0 could result. Thus, the flow rate of the working fluid should be reduced before the mode change. The control process is presented in a T-S diagram in Fig. 8(b). Before the mode change, reducing the flow rate decreases the expander inlet pressure and increases the superheating (process 1e20 in Fig. 8(b)). After the mode change, sufficient superheating could be maintained even though the expander inlet pressure is increased (process 20 e30 in Fig. 8(b)). When the operating mode of the PE-ORC system changes from low capacity to high capacity, the pressure at the expander inlet decreases, and the superheating could be larger than that before
211
the mode change (process 30 e20 in Fig. 8(b)). Thus, the vapor quality would not be an issue, and the sufficient superheating allows the expander inlet pressure to be increase for system performance (process 20 e1 in Fig. 8(b)). The control method in this case is to increase the expander inlet pressure by increasing the flow rate of working fluid after the mode change. 4.2. Test for operating mode changes Yun et al. [23] investigated the general behaviors of the total power output during operating mode change for PE-ORC using two identical expanders. The degree of superheat before the expander inlet is an important control parameter for preventing expander
Fig. 8. T-s diagram for operating mode change. (a) Total shaft power. (b) Cycle efficiency.
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E. Yun et al. / Energy 113 (2016) 204e214
damage, so this study focuses on the superheating variation when changing from high capacity mode to low capacity mode (from DUAL MODE to SINGLE MODE 1 and from SINGLE MODE 1 to SINGLE MODE 2). The rotational speeds of both expanders were made the same as in the performance characterization test. All tests were carried out with a fixed rotational speed of the working fluid pump for each case. A degree of superheat before operating mode changes was considered to be higher than 25 C based on the vapor quality drop after the mode change. The test results in Fig. 9 are superimposed on the performance curves from Fig. 6 and are summarized in Table 4. The results reveal good reproducibility of the performance curves. For both cases, the total shaft power and the cycle efficiency increased after each mode change, while the evaporative heat transfer rate decreased. The
Table 4 Data recorded during the dynamic tests. Case
Operating mode change
1
DUAL MODE to SINGLE MODE 1 SINGLE MODE 1 to SINGLE MODE 2
2
Q_ ev
_ ex;tot W
hc
(kW)
(kW)
(%)
(bar)
(kg/s)
(K)
20.84 19.22 10.44 9.406
1.21 1.56 0.52 0.73
5.88 7.92 4.76 7.35
6.75 10.65 5.96 8.99
0.07802 0.07382 0.03903 0.03591
26.38 11.43 28.32 11.96
Pex,in
m_ r
DTsup
evaporative heat transfer rate was slightly reduced by the reduced flow rate of working fluid due to the rapid increase of the pump outlet pressure. The meaningful increases in total shaft power are produced by the increased pressure ratio.
Fig. 9. Mode change test results superimposed on performance curves from Fig. 7. (a) Case 1 (b) Case 2.
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Fig. 10. Temporal variations in pressures over the expanders, mass flow rate of working fluid, superheating at expander inlet and evaporative heat transfer.
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Fig. 10 presents the temporal variations in inlet/outlet pressures, mass flow rate, degree of superheat and evaporative heat transfer rate during mode change tests. For both cases, there is a significant drop of more than 15 C in the superheating as the operating mode changes from high to low capacity. Thus, the mass flow rate of working fluid should be reduced appropriately for sufficient superheating before the operating mode change to maintain the state of expander inlet as superheated after the change. After the mode change, the mass flow rate of the working fluid decreases slightly due to the significant increase of the pump outlet pressure. This decrease depends on the pump performance and could be considered when determining the flow rate reduction for the operating mode change. While the enthalpy difference through the evaporator is almost constant during the mode changes, the evaporative heat transfer rate is slightly reduced after the operating mode changes. This depicts that the variation in evaporative heat transfer is strongly affected by the variation in mass flow rate of working fluid during the operating mode changes.
5. Conclusions This study experimentally investigated the performance characteristics and control methods for a PE-ORC system using two expanders with different capacities. A test bench was built with a PE-ORC loop and two different scroll expanders, an electrical water heating system, and an air-cooled chiller. The maximum number of operating modes (Nop,max) is 2n1 for n expanders in the PE-ORC. The number of operating modes was determined to be 3 due to the different expander capacities for DUAL MODE, SINGLE MODE 1, and SINGLE MODE 2. The results of the characterization test showed that the performance in DUAL MODE closely matched the average expander performance and could be used to design a PE-ORC system. The tested performance curves against the evaporative heat transfer rate showed three clearly separated operating regions with high efficiency and two possible switching points for the operating mode. The test results reveal that the proposed system could have a good heat recovery capability over a wide variation of heat input. Control strategies for the PE-ORC system were analytically and experimentally studied. The results showed that a significant drop in the superheating occurred after the operating mode changed from high to low capacity. The expander inlet pressure should thus be reduced appropriately for sufficient superheating before the operating mode change by decreasing the flow rate of working fluid. Control methods were proposed for PE-ORC operation to maintain the state of expander inlet as superheated and achieve high-performance heat recovery in PE-ORC system.
Acknowledgements This research was partially supported by Basic Science Research Program through the National Research Foundation of Korea (NRF) funded by the Ministry of Education (NRF-2013R1A1A2012173), and partially by the National Research Foundation of Korea (NRF) grant funded by the Korea Government (MSIP) through GCRC-SOP (No. 2011-0030013).
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