IMD

By Xudong Wang a, Yunho Hwangb,*, Reinhard Radermacherba
Air-Conditioning, Heating, and Refrigeration Institute, 2111 Wilson Boulevard, Suite 500, Arlington, VA 22201-3001, USA.

Centre for Environmental Energy Engineering, Department of Mechanical Engineering, University of Maryland, College Park, MD 20742, USA

Experimental results

1. Steady state cooling and heating tests
The system performance of the IHXC and the FTC in both cooling and heating modes was compared to the baseline system (Wangetal. 2008).The changes of the system capacity and the COP at different injection ratios are illustrated in Fig.

Fig. 4 – Comparison of the VI cycle and the baseline performance.
Overall, the IHXC and the FTC show a comparable performance improvement. The maximum cooling capacity gain is 15 with a 2 percent COP gain at the ambient condition of 46.1 °C. The maximum cooling COP improvement is around 2W 4 percent depending on the ambient conditions. The maximum heating capacity gain varies from 13 percent to 33 percent as the ambient temperature decreases from 16.7°C to 17.8 °C. The Improvement of the heating COP is more significant at the low ambient conditions than that at the high ambient conditions. The maximum COP improvement is 23 percent for the FTC at the ambient temperature of 17.8 °C.

The results show that this cycle option is more favourable for the high ambient cooling and low ambient heating applications (Wangetal.,2008).The overall comparison of the VI system capacities and the baseline conventional system capacities is illustrated in Fig. 5 by using the FTC as an example. It can be observed that the overall system cooling and heating capacities are significantly improved, and the load and capacity balance temper range are extended by applying the vapour-injection cycles to the conventional system. The conventional system design points for cooling and heating have been extended from 35°C and 5°C to 37°C and 8°C, respectively, with the assumption of the cooling and heating starting at 16.7°C.

Fig. 5 – Comparison of the VI cycle and the baseline capacities

1.2. Comparisons of the IHXC and the FTC
If perfect separation in the flash tank occurred and the injected refrigerant super heat at the out let of the internal heat exchanger was zero, then the ideal cycles of the IHXC and the FTC would be thermodynamically identical. However, there are applications of these two cycles show some differences

1.3 Operating range in terms of injection pressure
It can be observed from Fig.4 that the IHXC has a much wider range of the injection ratio than the FTC in terms of injection pressures. In practice, thermostatic expansion valve (TXV) control is normally applied at the injection line of the IHXC for simplicity. This gives a freedom of setting the injected refrigerant superheat. A small amount of superheat setting results in a large injection ratio and a high injection pressure, and vice versa. In such a case, the injected refrigerant leaving the internal heat exchanger is always superheated. The injection of saturated vapour (0K super heat setting) is difficult to achieve without causing TXV hunting (BeetonandPham,2003).On the other hand, the FTC utilizes the refrigerant vapour/liquid-phase separation to fulfill the cycle operation so that the injected vapour is saturated when leaving the flash tank. Hence, there is not much room to adjust the injection pressure at all when the compressor volume ratio and the system charge are fixed.

1.4. Impacts on condenser
Most of the systems in the previous research were equipped with liquid receivers at their condenser outlets such as the works by Winandy and Lebrun (2002) and Wang (2005).In those cases, the sub cooling at the condenser outlet can be considered approximately constant (Wang,2005). Unlike those systems, the tested heat pump system in this research does not have a receiver in its liquid line for the purpose of reducing the overall system cost and size. The change of the condensing pressures at different injection pressures is illustrated in Fig.6.

Fig.6– Effect of the intermediate pressure for vapor injection on condenser operating conditions.
As shown there, the vapour injection has little effect on the condensing pressure. This is mainly because the sub cooling of the system significantly decreases as soon as the vapour injection starts. This reduces the compressor head pressure and the liquid-phase heat transfer area of the condenser. The overall heat transfer coefficient of the condenser increases due to the increase of the two-phase heat transfer area.

1.5. Impacts on evaporator
Ideally, the IHXC can make the refrigerant enthalpy at the evaporator inlet as low as that of the FTC. However, the IHXC shows higher refrigerant enthalpy at the evaporator inlet than that of the FTC as shown in Fig.7a, due to the existence of the approach temperature between the refrigerant main stream and the injected stream at the internal heat exchanger in the IHXC. The FTC reduces the refrigerant enthalpy at the evaporator inlet by refrigerant phase separation. It is more efficient in reducing the enthalpy than the IHXC. Fig.7b indicates that the evaporating pressure of the IHXC shows a decreasing trend with increasing the injection pressure at each ambient condition.

Fig.7–Effect of the intermediate pressure for vapour injection on evaporator operating conditions.
However, the evaporating pressure of the FTC at each ambient condition does not show obvious change with varying the injection pressure. From the heat transfer point of view, the temperature difference between the refrigerant side and the air side has to increase with increasing the evaporator load. Hence, the evaporating temperature associated with the evaporating pressure has to decrease to make such temperature difference larger. The decrease of the evaporating pressure has a negative effect on the heat pump system. This results in a reduction of the refrigerant mass flow rate in the evaporator

1.6 Impacts on expansion device for evaporator
The IHXC has one-stage expansion from the system condensing pressure to the evaporating pressure. On the other hand, the FTC has two-stage expansion. The pressure drop across the TXV at the evaporator inlet is only from the system intermediate pressure to the evaporating pressure, and is much smaller than that of the IHXC. This has a significant impact to the FTC performance. The tested heat pump unit was originally equipped with a TXV rated for a nominal capacity of 11kW.The super heat setting of the TXV was about 4K. The TXV properly worked during the baseline and IHXC tests. However, it caused certain performance degradation of the FTC. Therefore, it was replaced by a larger TXV rated for a nominal capacity of 18kW during the FTC test.

As shown in Fig.8a, the capacity of the FTC with the 11-kWTXV is about 300W less at the ambient temperature of 35C as an example; and the compressor power, shown in Fig. 8b, is About 69 Whigher than that of the IHXC, when both cycles are operated at their respective maximum capacity points. This makes the COP of the FTC 4.4 percent less than the IHXC. The FTC with the 18-kW TXV, however, has comparable performance to the IHXC.

Fig. 8c shows the refrigerant superheat of the IHXC with the 11-kW TXV and the FTC with the 11-kW and the 18-kW TXVs at the evaporator outlet. The 11-kW TXV of the FTC has an excessive amount of superheat ranging from 10 to 14K at different intermediate pressure levels. This indicates that the 11-kW TXV has already at its fully opened position so that its opening cannot be any larger. As the refrigerant flow is restrained by this limited valve opening, the refrigerant is highly super heated in the evaporator. Moreover, the refrigerant density at the compressor suction line is reduced in the FTC with the 11-kW TVX because of the excessive amount of superheat, which results in a reduction of the refrigerant mass flow rate as shown in Fig. 8d. This phenomenon can happen not only to the FTC cooling application, but also to the high ambient FTC heating application where the system has comparable refrigerant mass flow rate at the evaporator side to the cooling application

Fig. 8 – Effect of the intermediate pressure for vapour injection on cycle performance and parameters.

1.7. Discussion on the control strategies of the vapour injection heat pump

Fig. 9a shows the cooling capacity of the IHXC under the different degrees of superheat of the injected vapour. The capacity increases with decreasing the injected vapour superheat. The point with lowest superheat setting (0.6K) at the ambient condition of 35 C shows a capacity reduction compared to the super heat setting of 9K at the same ambient condition. This is because the system lost the sub cooling at such a case; the refrigerant coming out of the condenser had a few vapour bubbles, which affects the reading of the refrigerant mass flow rate meter.

This point does not affect the overall trend. The COP of the IHXC is shown in Fig.9b.The COP increases with increasing the injected vapour super heat. For the IHXC heating application, the results in Fig. 4 show that the higher heating capacity and COP of the IHXC can be achieved as the more refrigerant is injected to the compressor. Hence the TXV superheat setting should be as small as possible for the IHXC heating application. In general, if there had to be a fixed super heating for both cooling and heating applications, 2W 3K would be a good balancing point.

The control of the FTC is somewhat more difficult than the IHXC. This is because the injected vapour of the FTC is obtained by phase separation. The refrigerant at saturated vapour condition is injected. The usage of a TXV as an expansion valve at the first stage of the FTC does not work in this case. The feasible option is to use an electronic Expansion valve (EEV) at the first stage of the FTC, since nearly 0K super heat control by EEV is feasible.

Fig. 9 – Effect of injected vapour superheat on the cooling performance

Conclusions
A two-stage heat pump system with a vapour-injected scroll compressor was developed, and tested. The FTC and the IHXC options of the two-stage vapour-injection system were investigated. The IHXC has a wider operating range of the injection pressure than that of the FTC due to its freedom of setting the injected refrigerant superheat at the injection port. Overall, the IHXC and the FTC show a comparable performance improvement as compared to the baseline system. The maximum cooling capacity gain is 15 percent with a 2 percent COP gain at the ambient temperature of 46.1°C. The maximum COP improvement is 2W 4 percent depending on the ambient conditions, which means that the vapour injection almost equally affects the capacity and the power consumption.

The heating capacity gain varies from 13 percent to 33 percent as the ambient temperature decreases from 16.7°C to 17.8°C. The maximum COP improvement, 23 percent, is achieved by the FTC at the ambient temperature of 17.8°C. For the IHXC, the simple and effective control is to use TXVs at the injection line and the main loop. For the FTC, it requires a larger TXV at the inlet of the evaporator than the one used in the IHXC.

Acknowledgment
We gratefully acknowledge the support of this effort from the sponsors of the Alternative Cooling Technologies and Applications Consortium and the Centre for Environmental Energy Engineering (CEEE) at the University of Maryland.

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