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Figure 3. Assumptions and limitations (underlined) on the ORC model. The exhaust gas was assumed to have the same composition as air. The coolant in the condenser was also assumed to be air, because air is the ultimate coolant on a vehicle. The isentropic efficiencies of both the pump and the expander were assumed to be 85%. The temperature difference between the working fluid and the exhaust gas at the outlet of the evaporator was assumed to be 10 °C (ΔT9-5 on Figure 3). The temperature difference at the pinch point in the evaporator was assumed to be 5 °C (ΔT8-3 on Figure 3). The temperature difference between the working fluid and the air at the outlet of the condenser was assumed to be 8 °C (ΔT1-10 on Figure 3). The temperature difference at the pinch point in the condenser was assumed to be 5 °C (ΔT7- 11 on Figure 3). These temperature differences reflect heat exchanger efficiencies, which should be decided in detail when design specifications are available. The temperature of the air before it enters the condenser was assumed to be 47 °C (T10 on Figure 3) according to [10]. The pressure of the exhaust gas in the evaporator was assumed to be 1.2 atmosphere pressure (1.2×101.325 kPa), and the pressure of the air in the condenser was assumed to be one atmosphere pressure. The mass flow rate of the exhaust was set to be unit mass flow rate, because this can provide us with a reference to the calculated mass flow rate of the working fluid and that of the condenser air without knowing the engine displacement and operating conditions. In addition, the heat exchange processes in the evaporator and the condenser were assumed to be externally adiabatic so that there is no heat loss. 11PDF Image | Working Fluid Selections in Organic Rankine Cycle ICE
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