PERFORMANCE AND CHARACTERISTICS OF COMPRESSOR/EXPANDER COMBINATION FOR CO 2 CYCLE M. FUKUTA, T. YANAGISAWA, S. NAKAYA (b), and Y. OGI Shizuoka University, 3-5-1 Johoku Hamamatsu, 432-8561, Japan Fax +81-53-478-158, e-mail tmmfuku@ipc.shizuoka.ac.jp (b) Shizuoka University, Graduate school of science and engineering ABSTRACT Since the air cooled CO 2 refrigeration cycle has a large throttling loss, an expander can improve the performance of the CO 2 refrigeration cycle by recovering the throttling loss. One way to utilize the recovered work is to drive an additional compressor by the expander. In such a case, it is effective to use an intercooler between a first-stage compressor and a second-stage compressor, which reduces the compression work of the second-stage compressor. A compressor/expander combination, in which the second-stage compressor is driven by the expander autonomously, is designed so that it is operated at a balance point of mass flow rate and shaft torque between the compressor and the expander. There are two balance points, however, and it is expected that the combined machine will be operated at the different condition from the original design point. The compressor performance was measured individually and it was expressed by a function of pressure increase by the compressor and rotational speed. The performance was used to calculate the balance point of the prototype compressor/expander combination. Although the performance of the combined expander is too low to drive the compressor, it ran autonomously with compressing the gas under the special experiment in which an additional flow is supplied to the expander. The operating point can be estimated using the performance maps of the compressor and the expander. 1. INTRODUCTION Transition from synthetic refrigerant to natural one is promoted in the refrigeration industry in recent years. Carbon dioxide (CO 2 ) is a potential candidate of the natural refrigerant because it has no flammability and no toxicity. The inherent performance of air cooled CO 2 cycles is lower than that of traditional cycles working with HFCs, and the efficiency of the CO 2 cycle should be improved. Since the CO 2 cycle has pretty large throttling loss at the expansion process, it is important to recover the loss in order to improve the performance of the CO 2 cycle. The use of an expander is one way to recover the throttling loss. Various types of the expander were proposed for the last decade. Stosic et al. (22) developed a twin screw expander combined with a compressor, Beak et al. (22) tested a piston-cylinder expansion device, Huff et al. (23) presented the performance of a scroll expander, Quack et al. (24) propose an integration of a three-stage expander in the series of their studies and Tφndell et al.(24) introduced a concept of an impulse type turbine. Authors (Fukuta et al., 2, 21, 23) have been studying the feasibility of a vane type expander, and the performance of the expander for the CO 2 cycle is being clarified theoretically and experimentally. How the recovered power by the expander is utilized is another issue to consider. In this study, the configuration in which a second-stage compressor is driven by the expander is examined. In such configuration, since the second-stage compressor and the expander is operated at the point that mass flow rate and power must be balanced between the compressor and the expander, individual
performance of the compressor and the expander influences the operating condition. Therefore, the performances of the second-stage compressor and the expander are examined respectively at first and then a prototype of vane type compressor/expander combined machine is designed, made and evaluated experimentally. 2. CHARACTERISTICS OF SYSTEM WITH EXPANDER 2.1. System configuration and performance There are many types of system configuration which integrate the expander (Hiwata et al., 23). Basically they are categorized into the following four types. (1) Direct drive type; the expander is connected to the main compressor and a motor with a common shaft. (2) Generator type; a generator is driven by the expander and the recovered power is used as electric power. (3) Low-side drive type; the compressor has a two-stage compression and the expander is connected to the first-stage auxiliary compressor. (4) High-side drive type; the compressor has the two-stage compression and the expander drives the second-stage compressor. In this study, the high-side drive type is considered because the second-stage compressor can be made very small, an arrangement of a connecting pipe is relatively flexible, less heat loss in the expander is expected and an intercooler between the first and the second compressors is available. Figure 1 is the P-h diagram for the cycle having the two-stage compression with the intercooler and an isentropic expansion process with the expander. It can be seen that the compression work at the second compressor is reduced by the intercooler. Figure 2 shows COP improvement by the expander and the intercooler versus intermediate pressure. The baseline of the COP improvement is COP of the cycle with neither the expander nor the intercooler. The conditions are; evaporating pressure is 4 MPa, heat rejection pressure is 1 MPa, heat rejection temperature of the intercooler and a gas cooler is 4 C, and efficiencies of the compressor and the expander are.7. It is assumed that the recovered power by the expander can be used as the part of compression power, not only at the second-stage compressor but also at the first-stage compressor. The COP is improved by 38 % by using the expander due to the reduction of the compression power and the increase of refrigeration capacity. When the intercooler is used, the maximum COP ratio becomes 1.51, because the second-stage compression power decreases. Since the discharge temperature for an isentropic compression becomes 4 C at 6.4 MPa, there is no improvement by the intercooler under the intermediate pressure of 6.4 MPa. Square symbols are the point that the second-stage compression power becomes equal to the expansion power, and the COP P[MPa] 14 1 6 expander and two-stage compression with 4 without COP ratio 1.6 1.5 1.4 with intercooler without intercooler 2 2 3 4 5 6 h[kj/kg] Fig.1 P-h diagram 1.3 6 7 8 9 P i [MPa] Fig. 2 COP improvement 1
becomes almost maximum under such condition when the intercooler is used. The second-stage compressor is driven autonomously by the expander under this condition. In this study, a prototype of the compressor/expander combination, in which the second-stage compressor is driven by the expander and operated independently from a main compressor, is developed and its performance is examined. The intercooler causes some pressure loss and consequently increases discharge pressure of the first-stage compressor. The influence of the pressure loss in the intercooler on the COP ratio is shown in Fig. 3. With increasing the pressure loss, the COP ratio decreases, and it is found that an acceptable, i.e. installing the intercooler is effective at least in energetic meaning, pressure loss is.6 MPa under the condition. 2.2. Operating condition The compressor/expander combination is operated under the condition that mass flow rates through the compressor and the expander are equal and shaft torques of the expander and the compressor are balanced. The compressor/expander combination designed to operate at the balance point shown in Fig.2 is considered, and its operating characteristics is examined. In this case, if both compressor and expander are a positive displacement type, the ratio between compressor suction volume and expander inlet volume is 1:.28 when both volumetric efficiencies are.8. Figure 4 shows intermediate pressure (suction pressure of the second-stage compressor) and pressure increase by the second-stage compressor when the heat rejection pressure changes under constant mass flow rate. The given conditions and assumptions are; the evaporating pressure is 4 MPa, the heat rejection temperature is 4 C, the intercooler is used, the total efficiencies of both the compressor and the expander are.7. The specific work of the compressor and the expander are plotted in Fig. 4(b) under the condition shown in Fig. 4. The expander work slightly decreases with the heat rejection pressure, i.e. the inlet pressure of the expander, under the constant inlet temperature of 4 C. While the compressor work has concave shape because of the pressure rise by the compressor. The point where the compressor 1 work is equal to the expander work is the balance point. As can be seen in Fig. 4(b), there are two P i balance points. The right one is an original design point, i.e. the heat rejection pressure is 1 MPa, 5 the evaporating pressure is 4 MPa and the intermediate pressure is 8. MPa. When the expander work is larger than the compressor work, the rotational speed increases and the Δ P COP ratio 1.6 1.5 1.4 1.3 with intercooler without intercooler.2.4.6 Pressure loss [MPa] Fig. 3 Influence of pressure loss at intercooler P[MPa] W[kJ/kg] 2 1 (b) Design point Wc We 8.5 9 9.5 1 1.5 11 P h [MPa] Fig. 4 Balance point
intermediate pressure decreases. Therefore, the left balance point seems to be stable but the right does not. As a result, it is expected that the compressor/expander combination is operated at the different point from the original design point. The operating condition at the left balance point is; the heat rejection pressure is 8.9 MPa and the intermediate pressure is 7. MPa. Such characteristic will be influenced by the efficiencies of the second-stage compressor and the expander, and has to be confirmed by an experiment. 3. PERFORMANCE OF SECOND-STAGE COMPRESSOR Since the second-stage compressor and the expander are operated at the point that flow rate and power is balanced between them, individual performance of both influences the operating condition. The total efficiency of a vane type expander we tested recently is about 6 % and the characteristics of the expander will be reported in another paper. Since the operating condition of the second-stage compressor varies widely as compared to the expander depending on the balance point and an operating condition of the first-stage compressor, the characteristics of the second-stage compressor is examined in this study. The compressor has the same structure as the compressor element in the compressor/expander combination except that shaft is not connected to the expander and goes out to the outside. The compressor is a vane type and its suction volume is 3.2 cm 3 /rev.. Its torque is measured by a torque meter inserted between the compressor and an inverter driven motor and mass flow rate is measured by a Coriolis mass flow meter. Since the pressure rise by the compressor in the combined machine will change according to the operating condition, the compressor performance is examined with changing the pressure difference. Figure 5 shows the volumetric efficiency versus the pressure difference with the rotational speed as a parameter. The volumetric efficiency decreases with an increase of the pressure rise and with the decrease of the rotational speed. Figure 5 (b) shows a value of the total efficiency divided by the volumetric efficiency, η t /η v. This value corresponds to a product of the indicated efficiency and the mechanical efficiency. When the pressure difference decreases, the efficiency decreases because mechanical loss increases relatively, and when the rotational speed increases the mechanical loss and indicated loss caused by flow restriction increases. These efficiencies can be expressed by functions of the pressure difference and the rotational speed. The functions are used to calculate the balance point of the prototype compressor/expander combination. 4. COMPRESSOR/EXPANDER COMBINED MACHINE ηv ηt/ηv 1.5 1.5 15rpm 2rpm 25rpm 3rpm 15rpm 2rpm 25rpm 3rpm The compressor/expander combined machine, both compressor and expander mechanisms are the vane type, was developed and its performance and characteristics were examined. The first prototype, however, was not operated autonomously because the expander was much smaller than that we tested independently and its performance was too (b).4.8 1.2 1.6 P[MPa] Fig. 5 Compressor performance
low to obtain the balance point for the mass flow rate between the compressor and the expander. We tried to verify the operating characteristics of the compressor/expander combination by supplying an additional flow rate to the expander. 4.1. Experimental setup Figure 6 shows an appearance of the compressor/expander combination developed. The compressor and the expander are vane type and both are connected with a common shaft in it. In order to measure the pressure change during the compression process, two piezo-electric pressure transducers are attached to the compressor side. The rotational speed is obtained from frequency of the pressure signal. The combined machine is connected to an experimental CO 2 refrigeration cycle. Discharge gas compressed by a main compressor passes through an intercooler. A part of the discharge gas leaving the intercooler is depressurized just for experimental purpose by a valve to the intermediate pressure and supplied to the second-stage compressor, i.e. the compressor of the combined machine. The pressure of the discharge gas from the second-stage compressor rises to the discharge pressure of the main compressor again and then joins the rest part of the discharge gas. The gas is supplied to the expander after passing through a gas-cooler and expands at the expander to the evaporating pressure. Consequently an additional flow is supplied to the expander as compared to the flow rate through the second-stage compressor. Note that since the second-stage compressor does not reduce the discharge pressure of the main compressor, the second-stage compressor does not have any beneficial work in this special experiment and this test is done only to verify the characteristics of the combined machine. The ratio of the flow rate through the second-stage compressor to that through the expander is changed to adjust the opening of the valve. The main compressor is scroll type for a residential CO 2 heat pump water heater. The gas-cooler, the intercooler and an evaporator are double tube type. Pressures and temperatures at each position in the cycle are measured by Bourdon tube pressure gauges and T-type thermocouples respectively. The mass flow rate through the second-stage compressor is measured by a Coriolis type mass flow meter. The mass flow rate through the expander is obtained using the same mass flow meter when the bypass valve is closed. The flow rate through the main compressor and the expander is assumed to be constant under the same operating condition of the main compressor. In the experiment, with keeping the operating condition of the main compressor, the mass flow rate through the second-stage compressor is changed by the bypass valve, and the rotational speed, pressures and temperatures at Expander Gas cooler 2 nd stage compressor Flowmeter Intercooler Evaporator Compressor Fig. 6 Compressor/expander combination Fig. 7 Experimental refrigeration cycle
the suction and the discharge of the second-stage compressor are measured. 4.2. Experimental results Figure 8 shows the rotational speed, and the suction and discharge pressures(b) of the second-stage compressor against the ratio of the mass flow rate through the second-stage compressor to that through the expander. Symbol shows the experimental result and lines are the calculated one explained later. As shown in the experimental result, the second-stage compressor can be driven by the expander and compresses the gas from 7.9 MPa to 8.7 MPa when the mass flow rate through the compressor is small, in other words, excess mass flow rate is supplied to the expander. When the ratio of the mass flow rates is smaller than.26 or larger than.68, the compression does not occur although the compressor/expander combination rotates. Figure 9 shows the calculated balance point of the mass flow rate and the shaft torque. The rotational speed, the intermediate pressure(b) and the shaft powers(c) of both the compressor and the expander are plotted against the heat rejection pressure for each flow rate ratio. In the theoretical calculation, the compressor performance shown in Fig. 5 is considered by approximating the performance by the function of the pressure difference and the rotational speed. On the other hand, the expander efficiency is set constant because the change of the expander performance is relatively small in the experimental range. When the compressor performance shown in Fig. 5 is taken into consideration, there is only one balance point of the shaft power for each flow rate ratio in most cases. The rotational speed and the pressure increase at the 4 balance point are shown in Fig. 8 with the solid or broken line. The calculated result has a 3 certain deviation from the experimental result. The reason why the balance point is not 2 estimated well is the fact that the combined compressor has a large clearance and it is 1 supposed that the performance of the combined N[rpm] N[rpm] P[MPa] 3 2 1 9.4 9 8.6 8.2 7.8 (b) P h P i Experiment Calculation Calculation Experiment.2.3.4.5.6.7 G c /G Fig. 8 Rotational speed and pressure condition of combined machine P h P i Pi[MPa] L[W] 1 5 16 8 (b) L e P h Gc/G=.67 Gc/G=.55 Gc/G=.41 Gc/G=.26 Gc/G=.67 Gc/G=.55 Gc/G=.41 Gc/G=.26 7 8 9 1 11 P h [MPa] Fig. 9 Estimation of operating condition (c)
compressor is lower than that shown in Fig. 5. Although the calculated result does not agree well with the experimental result, it has a similar tendency. It is expected that the compressor/expander combination is operated as predicted by the balance calculation if exact performance and characteristics of the combined compressor and expander is clarified. Since the compressor performance is particularly influenced by the pressure condition and the rotational speed, mapping the performance against the pressure condition and the rotational speed is important to calculate the operating point. 5. CONCLUSIONS In order to improve the cycle performance of the CO 2 refrigeration cycle, recovery of the throttling loss is an important issue. It was shown that a two-stage compression with an intercooler can improve the cycle performance as well as the power recovery by an expander. Under the condition that evaporating pressure is 4 MPa, heat rejection pressure is 1 MPa, heat rejection temperature of the intercooler and a gas cooler is 4 C, and efficiencies of the compressor and the expander are.7, the COP will be improved by 51 % using the expander and the intercooler as compared with a normal CO 2 cycle. A compressor/expander combination, in which the second-stage compressor is driven by the expander autonomously, was designed so that it is operated at a balance point of mass flow rate and shaft torque between the compressor and the expander. There are two balance points, however, and there is a possibility that the combined machine will be operated at a different point from the original design point. The compressor performance is measured individually and it is expressed by a function of pressure increase by the compressor and rotational speed. The performance is used to calculate the balance point of a prototype compressor/expander combination. Although the performance of the combined expander is too low to drive the compressor, it ran autonomously with compressing the gas from 7.9 MPa to 8.7 MPa under the special experiment in which an additional flow is supplied to the expander. The operating point can be estimated from the balance point of mass flow rate and shaft torque between the compressor and the expander. Since the compressor performance is particularly influenced by the pressure condition and the rotational speed, mapping the performance against the pressure condition and the rotational speed is important to calculate the operating point. REFERENCES Beak JS, Groll EA, Lawless PB. 22, Development of a piston-cylinder expansion device for the transcritical carbon dioxide cycle, Proc. of 9th Int. Refrig. and Air Conditioning Conf. at Purdue: R11-8. Fukuta M, Radermacher R, Lindsay D, Yanagisawa T. 2, Performance of vane compressor for CO 2 cycle, Proc. of the 4th IIR-Gustav Lorentzen Conf. on Natural Working Fluids at Purdue: 339. Fukuta M, Yanagisawa T, Ogi Y, Radermacher R. 21, Cycle performance of CO 2 cycle with vane compressor-expander combination: Trans. on Compressors and their Systems: 315-324. Fukuta M, Yanagisawa T, Radermacher R. 23, Performance Prediction of Vane Type Expander for CO 2 Cycle, Proc. of the 21st IIR Int. Congr. of Refrig.: ICR251. Hiwata A, Lida N, Sawai K. 23, A study of cycle performance improvement with expander-compressor in air-conditioning systems: Proc.of the IMechE Conf. Trans. on Compressors and their Systems: 339.
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