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as the advanced design. The sizing of the SCO2 cycle is performed in such a manner that maximum efficiency is reached at minimum cost since large components, in particular heat exchangers lead to higher cycle efficiencies due to lower pressure drops, but increase cost. A computer code has been developed and the optimization performed as follows (see Refs. 10 and 13 for more details): 1. Main compressor inlet temperature was selected as a compromise between the cycle efficiency and prevention of operation below the critical temperature of CO2 because of uncertainties with compressor behavior below this point. The closer the temperature to the critical tempe- rature, the higher the efficiency, hence 32°C was selected. A CEA study, reported in Section 3.5, explored the region below the critical temperature to further increase efficiency. 2. Main compressor outlet pressure was selected based on cycle efficiency sensitivity to pressure and economic considerations. The analyses showed a sharp efficiency decrease for operating pressures below 20 MPa but very small increase for pressures above 20 MPa. Because the economic penalty for pressure above 20MPa would begin to dominate benefits of the small efficiency increase and because it is desirable to remain within the experience base of supercritical steam cycle pressure, 20MPa to 25MPa was selected as the operating pressure of the SCO2 cycle. The CEA study, reported in Section 3.5, explored the impact of a maximum operating pressure of 25 MPa on the cycle efficiency. 3. Total volume of the cycle heat exchangers (recuperators and precooler) has been selected to yield minimum specific capital cost ($/kWe). This volume was found to be about 120m3 for a target plant capital cost of 1000$/kWe 4. Given the optimum total volume, the individual heat exchangers were optimized to yield the highest cycle efficiency. This involved optimum split of the total volume among individual heat exchangers and identification of the optimum length and face area of each heat exchanger. The optimum volume apportionment between high- and low- temperature recuperator and precooler was found to be 53, 46, and 21m3 with corresponding active lengths of 1.75, 2.05 and 1.1 m, respectively [13]. All heat exchangers were of HEATRICTM’s printed circuit type with straight channels, semicircular channel diameter of 2mm, plate thickness of 1.5mm and pitch of 2.4mm and fully countercurrent flow. Heat exchanger sizes, operating pressure and main compressor inlet temperatures are the same for both the basic and advanced designs. Operating conditions of both designs for compressor inlet temperature above the critical value are summarized in Table 1 and the statepoints for each design are given in Table 2. The net efficiency, Net, was evaluated from the following equation: where : M is the coupling mechanical efficiency (99%) (one coupling between each turbine or compressor body, i.e. 3 couplings) is the generator efficiency (98%) SY is the switchyard loss (0.5%) P accounts for system parasitic losses (2%) WNET is the net turbine work WCT is the total work of compressors WCR is the work of the recompressing compressor WCM is the work of the main compressor WP is the pumping power for water cooling Qth is reactor thermal power (added heat rate) Turbomachinery efficiencies used in the cycle calculation were those obtained from the detailed design at MIT of the turbine and compressors. A fixed pressure drop of the heat source of 130kPa was used in the cycle optimization. In the indirect cycle, this pressure drop depends on the design of the IHX. As mentioned in the introduction, to avoid reoptimization of the entire cycle, the dependence of cycle efficiency on heat source pressure drop was calculated and used in performance evaluation of the indirect GFR cycle presented in Section 3. This dependence is fairly linear and can be accurately described by the equation = 0-0.002 p IHX, where 0 is cycle efficiency for zero IHX pressure, pIHX= 0kPa. The attractive feature of the SCO2 cycle is extremely compact turbomachinery. The dimensions, number of stages and efficiencies of the turbine [9] and compressors [14] are summarized in Table 3. It can be noted that all of the blading sections of the turbomachinery for a 300MWe unit can fit in a home-size refrigerator. This is because of small volumetric flow rates due to high compressor inlet and outlet pressures. It is of interest to compare the net efficiencies of the above SCO2 cycles to the efficiency of a direct helium Brayton cycle, such as the GT-MHR. To make a consistent comparison the net efficiency of a helium cycle with 1 intercooler and 2 compressors and turbine inlet temperature of 850°C (same as GT-MHR) was calculated, using the same procedure, and found to be 46.3%*. This is 2.5% higher than the net efficiency of the basic design at 550°C core outlet temperature and about 1.5% less than the advanced design at core outlet temperature of 650°C. HEJZLARetal., AssessmentofGasCooledFastReactorwithIndirectSupercriticalCO2Cycle G * GT-MHR net efficiency is reported as 48% [15]. This efficiency was reproduced by our calculations when neglecting water pumping power and parasitic system losses. Higher GT-MHR efficiency may indicate that the parasitic system losses assumed in our calculations may be conservatively high. NUCLEAR ENGINEERING AND TECHNOLOGY, VOL.38 NO.2 SPECIAL ISSUE ON ICAPP ‘05 111PDF Image | GAS COOLED FAST REACTOR WITH INDIRECT SUPERCRITICAL CO2
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