Since the refrigerant boils outside the tubes carrying water, the refrigerant side heat-transfer coefficient varies along its length, and the procedure explained in the “Shell and tube heat exchanger with phase change in the shell-side fluid” in Section 4.1.4 has been adopted to predict its performance. 4.15 shows the schematic diagram of the chiller. The fins are 0.379 mm thick and spaced 0.958 mm apart from each other (19 fins per inch of tube length). The evaporator of this plant is a two-pass flooded liquid chiller, each pass having 114 integrally finned tubes each of 3.64 m effective length, 13.8 mm ID, 17.04 mm OD at fin root and 19.05 mm diameter over the fins. Dhar, in Thermal System Design and Simulation, 2017 Evaporator Therefore, the optimized cycle power savings is due mainly to the optimized mass distributions and pressure levels. In order to know how much power savings were obtained due to lower condensing temperature, 40☌ was applied on the baseline model, which resulted in 0.817 MW power reductions from the total baseline cycle power consumption. The other split ratios, x 11, x 13, and x 14, were adjusted by the optimizer to meet the change in the total mass flow rate. Split ratios, x 8 and x 9, are reduced so that an optimized amount of refrigerant is provided to low temperature heat exchangers at –19☌ and –33☌, respectively, which has low expansion pressure that requires more compression power. The optimized propane cycle has a higher refrigerant mass flow rate, maximum subcooling, and slightly lower overall compression ratio than those of the baseline cycle as shown in Table 5-12. The optimized propane cycle has a power consumption of 37.15 MW, which is 15.98% less than the baseline power consumption. Figure 5-8 and the LMTD values show that the cold curve in the optimized cycle is closer to the hot curve than the baseline cycle, which means more efficient heat transfer or less entropy generation in the heat exchanger. The SWHX has two sections with log mean temperature difference (LMTD) for the optimized cycle of 5.24☌ and 4.91☌, whereas the LMTDs of the baseline cycle are 7.12☌ and 5.17☌. A plot of the cooling curves for the optimized and baseline MCR cycles is shown in Figure 5-8. On the other hand, propane has the highest boiling temperature, which increases the refrigeration capacity of the refrigerant. Since nitrogen has the lowest boiling temperature among the constituents, it lowers the lowest refrigeration temperature. Nitrogen and propane mass fraction increased while the methane and ethane decreased. The optimized MCR cycle has lower refrigerant mass flow rate and lower overall compression ratio than those of the baseline cycle as shown in Table 5-11. The optimized MCR cycle has a power consumption of 63.63 MW, which is 4.48% less than the baseline power consumption.
The defrosting duration for the ASHP unit was the same as that of Circuit 3. 4.20B and C were derived following this assumption. The experiment conditions illustrated in Figs. When the tube surface temperature at the exit of a refrigerant circuit reached 24☌, the defrosting operation on that circuit was considered ended. The energy consumption on the compressor would decrease 33% as a result of a compressor speed reduction by 33%. In Study Case 3, as the modulating valve on Circuit 1 was closed, the refrigerant mass flow in Circuit 2 and Circuit 3 remained unchanged as a result of compressor speed reduction by 33%. As a result, the refrigerant mass flow rates to Circuit 2 and Circuit 3 were each increased by 50%, as illustrated in Fig. In Study Case 2, the total refrigerant mass flow was evenly distributed to the other two refrigerant circuits during defrosting after the modulating valve on Circuit 1 was closed. Assumed refrigerant mass flow rate in the three circuits during defrosting in Study Case 1.