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294 As shown in Figure 8, the processes of evaporation and liquid heating are the most critical heat exchange pro- 295 cesses to the implementing of the RC and ORC configurations in a vehicle. The volume requirements for the RC 296 configuration with dT=10◦C (1i and 3i) could be too large for the implementation in a HDD engine. The ORC config- 297 uration seems to be the solution with less technical restrictions from the point of view of volume requirements for the 298 heat exchangers. In steps (5) and (6) of Figure 1, an iterative process is specified; where dT has to be progressively 299 increased in order to fulfill the size due to packaging constraints in the vehicle. As we do not have any particular 300 vehicle with a particular packaging objective, we are going to skip such iterative process and straight forward propose 301 an increment of dT from 10◦C to 20◦C in order to exemplify its effect on reduction of bsfc. The configuration with 302 high temperature sources suffers a significant penalty of bsfc in almost all the working points due to increase of dT. 303 Figure 9 shows the results of this study in the configuration with high temperature sources (case B) and ideal cycle 304 assumption. This figure shows a reduction of bsfc improvements in the range of 8% to 4% for the ORC configuration, 305 and of 11% to 2% for the RC configuration. 306 Respect the condenser size, the new space requirement is practically negligible due to his place in the grille and 307 the high considered vehicle speed. The effect of dT=20◦C in the sizing of the heat exchanger is evaluated for the operating points with higher reduc- 309 tion of bsfc for the RC and ORC configuration, i.e. 1800 rpm and 100% load for RC, and 1200 rpm and 50% load for 310 RC. These points are called as 5i and 6i respectively, as shown in Table 4. The results of these sizing problems are 311 shown in Figure 10. For both configurations with high temperature sources, a dT=20◦C has implied a 50 % reduction 312 of the heat exchanger volume. Thus the volume requirements with dT=20◦C are more acceptable. According to these 313 results, the ORC with the high temperature sources could be the solution easier to implement in a HDD engine taking 314 installation constraints into account. 315 The configuration with all sources (case A) practically does not produce a net power output. It is result of the low 316 temperature difference imposed between evaporator and condenser in order to recover the waste heat of cooling water. 317 This temperature difference is so low that it doesn’t allow increasing the minimum temperature difference in the heat 318 exchangers (dT ). 319 The condenser size in these two configuration is also negligible compared to the other evaporator and pre-heater. 320 3.7. Influences of real expander and pump 321 In this section, the effects of internal irreversibilities in the cycle output power are evaluated. These irreversibilities 322 are considered only in the expander and the pump. The pump model is simulated with a isentropic compression 323 efficiency of 80% [6]. The isentropic expansion efficiency of the turbine is determined using the turbine chart which 324 is referenced in [24]. In this chart, the optimal expander and the isentropic efficiency can be obtained from two 325 dimensionless numbers: the Load Number and the Flow Number. These coefficients are calculated though the inlet 326 and outlet conditions of the expander machine [25]. 11 308PDF Image | Methodology to design a bottoming Rankine cycle
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