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233 The main differences were observed in the RC implementation. In this configuration with the high temperature 234 sources, the cooling water temperature does not imply a restriction in the evaporation temperature. For this reason, 235 higher evaporation and superheating temperatures as shown in Table 3 can be considered for the optimal cycle, for 236 instance, at 1800 rpm and 100 % load the superheating temperatures is close to 500 ◦C. Consequently, the reduction 237 of bfsc ranged from 10% to 14% at points of high thermal level (high speed and load), i.e. from 1200 to 1800 rpm and 238 100% load. The highest reduction of bsfc was achieved at 1800 rpm and 100% load. 239 3.6. Sizing heat exchangers 240 The technical implementation of a bottoming cycle in a HDD engine is not only based on thermodynamic de- 241 sign criteria. Also, the required space, weight and cost of installation are important criteria for the viability of the 242 considered solution. Thus, the sizing and the materials of the heat exchangers for the power cycles proposed in the 243 theoretical study are calculated. This involves the selection of heat exchanger construction type, flow paths, physical 244 size to meet the specified heat transfer and pressure drops within specified constraints [42]. For the sizing of the heat 245 exchangers, the following criteria have been imposed: 246 • 247 • 248 249 • 250 251 252 253 254 255 256 257 258 259 260 261 Aluminum was selected as the heat exchanger material. The heat exchangers were designed so that the pressure drop is less than 2%, in order to be consistent with the hypotheses considered in the previous step. Shell&tube heat exchangers have been chosen for the gas-liquid and liquid-liquid heat transfer. For the rating problem, the step-by-step developed by Delaware-Bell et al. [43] has been followed. An E-11 shell flow ar- rangement with a square pitch configuration is selected for all the shell& tube exchangers [44]. The number of passes and the log mean temperature difference correction factor F have been calculated using the Domingo’s correlation [45]. The Dittus-Boelter’s correlation for the heat transfer coefficient and Blasius/McAdams equa- tions for the friction factor were used for the calculation in the tube side [46]. The ideal heat transfer coefficient for the shell side was calculated by: jCpGs(Φs)n hid = 2 (4) Pr3 where the term j is the ideal Coulburn factor for the shell side and can be determined form the appropriate Bell-Delaware curve for the tube and pitch [43], Cp is the average fluid heat capacity at constant pressure along the heat transfer process, G s is the shell side mass flow velocity, the term (Φns ) is the viscosity correction factor, with accounts for the viscosity gradients along the tube wall versus the viscosity at the bulk mean temperature of the fluid and Pr is Prandtl number. The standard definition of this variable is given by Equation 5 taking into account the different temperature criteria for the gases and liquids being cooled or heated [47]: 9PDF Image | Methodology to design a bottoming Rankine cycle
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