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In the case A with all the heat sources and dT = 10◦C an expansion ratio of 3 and 2 for RC and ORC respectively 328 is produced in the expander machine at all the operating points. The reason for these low expansion ratios is the low 329 evaporation temperature imposed, in order to recover the cooling water heat source. It allows an expansion process 330 completed with a single expander machine, obtaining an acceptable isentropic efficiency of 70%. Regarding the type 331 of expander, the chosen solution for RC is a reciprocating piston expander. In the ORC case, the expander can be a 332 radial or axial turbine. 333 On the other hand, in the case B with high temperature sources (EGR, exhaust gases and aftercooler), a high 334 expansion ratio is produced in the expander as shown in Figure 11. In the ORC results, the expansion ratio is higher 335 than in case A (6 vs 2). The optimal solution for this configuration using the turbine chart referenced in [24] is still an 336 axial or radial turbine with an isentropic efficiency of 70%. On the other hand, the expansion ratio in RC configuration 337 (25 with high temperature sources vs 3 with all sources) is too large to be efficiently implemented in a single stage 338 expansion. Thus, the optimal expander is a two-stage reciprocating piston expander with 70% and 60% of isentropic 339 efficiency for the first and second stage respectively. 340 The irreversibilities of the expander machine and the pump are imposed in the studied configurations. 341 Figure 12 shows the effects of the real compression and expansion processes in the four studied cases with 342 dT=10◦C. In the configuration with all sources, the irreversibilities of the expander machine are less critical, be- 343 cause the low expansion ratio provides the possibility of using a single stage expander. Now, the highest reduction 344 obtained in bsfc for the case A is limited to 10%, at 1200 rpm and 25% load both for RC and ORC solution. An even 345 greater impact on bsfc is obtained in the case B configuration with a maximum reduction of bsfc above 10% at 1800 346 rpm and 100 % load and 7% at 1200 rpm and 25% load for RC and ORC respectively. Figure 13 shows the effects of the expander and pump irreversibilities with dT=20◦C in the case B. The highest 348 reductioninthiscaseisabove8%at1800rpmand100%loadforRCand5%at1200rpmand25%loadforORC 349 solution. 350 The effect of the pump and expander irreversibilities in the heat exchanger inlet temperature is almost negligible. 351 For this reason, the volume requirements were considered equal to those in the ideal cycle study. 352 4. Summary 353 Table 6 is a summary showing the highlighted numerical results of the illustrative example exposed through this 354 paper. The table shows the performance and requirement at optimum solution for each operating point. Taking into 355 account the ideal cycle assumption, dT = 10◦C and the configuration with all the sources (A), the bottoming cycle 356 reaches a maximum reduction of bsfc near 15% and 14%, for RC and ORC respectively (cases 1i and 2i). The case 357 B, with only the high temperature heat sources, ideal cycle and dT = 10◦C can achieve a reduction near 14 % and 358 10% for RC and ORC respectively (cases 3i and 4i). The RC configurations need a high heat exchanger surface and 12 327 347PDF Image | Methodology to design a bottoming Rankine cycle
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