## 3 Results
### Sharing a Common Shaft
Furthermore, the evaporation temperature becomes higher, explaining why the cooling capacity becomes larger for a higher condensation temperature, although the power provided by the ORC expander decreases. Figure 12 shows the contribution of each component on the exergy destruction under different condensation temperatures in the condenser. It is evident that under the operating conditions considered, the evaporator in the ORC power cycle contributes the most significant part of the total exergy loss in the system. This result can be attributed to the significant irreversibility associated with heat transfer across the large temperature differences in the evaporator. Figure 10: Effect of condensation temperature in both condensers on evaporation for the case of sharing a common shaft. Figure 9: Effect of ORC evaporation pressure on cooling capacity for the case of sharing a common shaft. Figure 11: Effect of ORC condensation temperature on cooling capacity for the case of sharing a common shaft. Figure 12: _Cont._  
In the VCC subsystem, the irreversibility across the compressor is the maximum under most working conditions. However, the exergy destruction in the compressor is higher than that when the condensation temperature is lower than 309.65 K and Peva = 1100 kPa. This suggests that in order to improve the performance of the proposed system, special attention should be paid to reducing the irreversibility that exists in these components during machine design. The irreversibility of the compressor and expander essentially depends on their isentropic efficiency; thus, their proper design can reduce this irreversibility. In particular, in order to reduce the irreversibility of the condenser and evaporator, they should be designed in such a way that the temperature difference between the fluids can be maintained as small as possible. Figure 13 shows the variation of exergy efficiency when the condensation temperature varies. Regarding the exergy input, it consistes of heating exergy of the exhaust gas and pump power input. The temperature of exhaust gas \(T_{6}\) calculated by using the PPTD method is lower than 393 K, but it is set to be a constant of 393 K to avoid the possible corrosion problems. Therefore, the heat input is the same, although the condensation temperature changed under considered conditions. Moreover, less pump power is required for a rising condensation temperature. Figure 12: Contribution of each component on exergy destruction under different condensation pressures. Regarding the useful exergy output, the cooling exergy is found to be reduced at its higher VCC evaporation temperature, explaining the reason as to why the exergy efficiency declines as the condensation temperature increases, except at \(P_{eva}\) = 1100 kPa. For \(P_{eva}\) = 1100 kPa, the exergy destruction of the expansion valve decreases sharply when the condensation temperature increases from 309.2 to 309.7 K (see Figure 12d), and a valley value at 309.7 K appeared, explaining why there is the maximum exergy efficiency under this condition. Furthermore, it can also be observed that a higher exergy efficiency could be obtained when the ORC power systemis operated at a higher evaporation pressure. This is because the evaporation temperature in the VCC is much lower, as shown in Figure 9.  
Table 1 summarises the optimal working conditions when the required cold air temperature is not higher than 263 K for a given heat sink temperature of 298 K. It can be understood that the belt drive transmission unit could result in the maximum cooling output of 9823 kW. Unfortunately, its ORC evaporation is 3500 kPa, which is much higher than that of the transmission unit sharing a common shaft. It is obvious that a thermo-economic evaluation is needed to decide which transmission unit is the best by making a trade-off between the thermodynamic performance and the compactness/reliability of the ORC-VCC combined system.