# Operational Optimisation of a Non-Recuperative 1-kWe Organic Rankine Cycle (ORC) Engine Prototype
## 4 Steady-State ORC Performance
### Part-Load Performance Maps
#### 4.2.1 Thermodynamic Analysis of the Optimum Configurations  
Control parameters were identified to maximise simultaneously the thermal and exergy efficiencies at various heat-source conditions. This section aims at further analysing the thermodynamic losses in these optimum configurations, both from a first- and second-law point of view. The steady operation and working-fluid states at key locations of the ORC engine are presented in Figure 5 that shows the temperature-entropy (\(T\)-s) diagrams of the optimum (maximising efficiency) configurations, as measured for heat-source temperatures of 120 and 140 \({}^{\circ}\)C. The working fluid circulated by the pump enters the evaporator in subcooled liquid state (1). The heat transferred from the thermal oil in the evaporator preheats the liquid-phase fluid (1-2), which then undergoes an isothermal evaporation (2-3) and is superheated in the last section of the heat exchanger (3-4). The heat addition (1-4) is a nearly-isobaric process occurring at \(9.6\pm 0.06\) bar and \(11.3\pm 0.06\) bar for heat-source temperatures of 120 and 140 \({}^{\circ}\)C, respectively. Thermal losses in the pipes lower the vapour superheat at the entrance of the expander (5), where the refrigerant expands (5-6), while the work done on the scroll shaft is evacuated in the resistive load bank. The working fluid is finally condensed (6-8) down to a subcooled liquid state before being sucked back into the pump. At the scale of the \(T\)-s diagram, the values of the temperature and specific entropy of the fluid at the pump inlet and outlet cannot be distinguished. Point (1) thus refers to both states. The exergy destroyed in the processes undergone by the working fluid appear graphically in the \(T\)-s diagrams. First, the vapour expansion process (5-6) is a non-isentropic process, due both to the thermal losses to the surroundings and to the irreversibility of heat- and mass-transfer occurring within the device. The performance of the scroll expander is further analysed in Section 4.2.2. The mismatch observed between the temperature profiles of the working fluid and heat-source and cooling water streams in the heat exchangers is another source of exergy loss. Indeed, theoretically, the temperature difference between the working fluid and the heating oil could be used to drive a Carnot heat engine. The area contained between the curve (1-4) and the heat-source profile (as well as between (6-8) and the heat-sink profile) is directly proportional to the exergy destruction in the evaporator (or condenser). From Figure 5, it is observed that the exergy destroyed in the evaporator: (i) is larger than that  
Figure 5: Temperature-entropy diagrams presenting the steady operation of the ORC engine with efficiency-optimum operating conditions for heat-source temperatures of: (**a**) 120 and (**b**) 140 \({}^{\circ}\)C. Working-fluid state reported at key locations of the facility: (1) pump inlet or condenser outlet, (2) evaporator saturated liquid, (3) evaporator saturated vapour, (4) evaporator exhaust, (5) expander inlet, (6) expander outlet, (7) condenser saturated vapour, (8) condenser saturated liquid. destroyed in the condenser and (ii) increases significantly with the heat-source temperature, while that destroyed in the condenser remains fairly constant between the two diagrams plotted. The contributions of the different components of the ORC engine to the overall exergy destruction is presented using pie charts in Figure 6. The latter presents the share of the condenser, evaporator, expander and pump in the thermodynamic losses throughout the cycle, for the three heat-source temperatures investigated, as taken from the efficiency-optimum configurations listed in Table 2. It is first observed that the losses incurred in the pump represents a constantly small amount compared to those in other components.