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both expander inlet pressure and temperature increased. Although the expander power output was primarily due to an increase of the mass flow rate, the change of pressure and temperature at expander inlet had secondary and beneficial effects. In addition, as illustrated in Fig. 16, the isentropic efficiency of the expander changes with the expander inlet and outlet pressure ratio. It also plays a role in determining the actual power output. Unlike modeling which can specify the change of one parameter at a time, it is not unusual that changing one parameter in a physical system affects several others during testing. As indicated in Fig. 16, the expander isentropic efficiency approaches 84% at a pressure ratio around 3.3 and decreases gradually as the pressure ratio is increased. The heat transfer effectiveness of the recuperator was evaluated based on the collected data at both the liquid and vapor sides. If the device was perfectly insulated and allowed to reach steady state, the heat transfer rates on both sides would be expected to be the same. The values are plotted in Fig. 17 with the horizontal axis showing effectiveness based on the liquid side and the vertical axis showing effectiveness based on the vapor side. According to the plot, effectiveness was consistently between 70 and 80 percent based on the enthalpy change on the liquid side, while effectiveness was on average 9% higher based on the vapor side. Although the recuperator was insulated with 0.5 inches thick melamine foam, it appears that the performance of the recuperator may be degraded by approximately 10% by heat loss during operation. This can be attributed to the significant amount of air flow inside the demo unit described in the previous section (draw-through air). As also shown in Fig. 15, the effectiveness of the recuperator increases with the mass flow rate, indicating heat loss became a smaller factor as fluid heat capacity increased. In other words, the thermal resistance associated with the heat loss became relatively larger as the mass flow rate of the power cycle increased. The power cycle efficiencies at various pump pressures are plotted in Fig. 18. For most of the cases where pressure was between 1,700 kPa and 2,000 kPa, the 1st law efficiency was around 10% and 2nd law efficiency was around 30%. Lower pump pressure seems to affect both 1st and 2nd law efficiencies in a negative way, which confirms what was obtained in the thermodynamic analysis – the higher the pump pressure, the higher the fluid saturation temperature in the boiler. This not only raises the heat input temperature, which improves the 1st law efficiency, but also reduces the temperature difference in the boiler which reduces entropy generation. Because the pump outlet pressure affects other important system parameters, higher pump pressure could also cause lower cycle performance due to poor isentropic efficiency of the expander at high pressure ratios, as illustrated in Fig. 16. These parameters include the pressure ratio at the expander inlet and outlet, and fluid superheat coming out of the boiler. Although the power cycle of the system performed relatively well during the tests, the performance of the cooling cycle was less than expected. Figure 19 shows the cooling capacity achieved, cooling cycle COP, and overall system COP. As illustrated, the cooling capacity increased with expander power as expected. However, both the cooling COP and overall system COP decreased as the expander power was increased. The direct cause for this was that the cooling capacity did not increase correspondingly with the power generated by the expander. This could be an inherent feature of the vapor compression cooling cycle; i.e. capacity is a function of several factors, not just power into the compressor alone. In general, 21PDF Image | Performance of a Combined Organic Rankine Cycle
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