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102 7 Waste heat recovery of large scale reciprocating engine 15 10 5 0 toluene n−pentane R245fa cyclohexane 80 100 120 140 o T , [ C] CA,out toluene n−pentane R245fa cyclohexane 15 10 5 0 (a) 80 100 120 140 o T , [ C] CA,out (b) 15 10 5 0 toluene n−pentane R245fa cyclohexane 80 100 120 140 o T , [ C] CA,out (c) Figure 7.5: The effect of charge air temperature at the outlet of the evaporator on the ratio of net power output to charge air mass flow rate for different organic working fluids with charge air inlet temperatures of (a) 180 ◦C, (b) 200 ◦C, and (c) 220 ◦C. needed in the charge air heat utilization compared to the exhaust gas heat utilization, leading to economically more unfeasible processes. The large heat transfer areas in CA utilization can be mainly explained by the smaller temperature differences between the heat source and the working fluid in the evaporator, as well as by the lower cycle effi- ciency when compared to the EG utilization, due to the lower evaporation temperatures. It should be noted that the charge air outlet temperatures corresponding to the highest power outputs of ORC are relatively high compared to typical charge air temperatures after the engine CAC. Thus, an additional cooling for charge air would be required after the ORC evaporator. This heat could be used for preheating the ORC working fluid thus replacing the recuperator or cooled down with the engine cooling loop. Also the pressure losses in charge air should be minimized because the pressure loss has a negative effect on P /q , [kW/kg/s] e,net m,CA P /q , [kW/kg/s] e,net m,CA P /q , [kW/kg/s] e,net m,CAPDF Image | WORKING FLUID SELECTION AND DESIGN OF SMALL-SCALE WASTE HEAT RECOVERY SYSTEMS BASED ON ORGANIC RANKINE CYCLES
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