Operation and Analysis of a Supercritical CO2 Brayton Cycle

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Operation and Analysis of a Supercritical CO2 Brayton Cycle ( operation-and-analysis-supercritical-co2-brayton-cycle )

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gas booster pump and the pump efficiency. The maximum Haskel pump inlet temperature is estimated to be 320 K, but could be substantially different depending on the amount of cooling provided to the housing and whether or not the rotor cavity is heated by magnetic losses. For the purposes of determining the Haskel pump pumping power, the pump inlet temperature is bounded by the lowest inlet temperature available which is 295 K (approximately the cooling water inlet T) and by our estimate for the highest temperature which is 320 K. The pump efficiency is estimated to be 85%. The pumping powers that were estimated for our design and scoping calculations, prior to operating the turbomachinery, are shown in the green row of Table 5.4. For a Haskel pump inlet temperatures of 320 K and for the largest leakage (choked flow conditions) and at an upstream pressure of 13,842 kPa, the pumping power was 12.24 kW. If the upstream pressure were 7,700 kPa, the pumping power would be 4.5 kW at 320 K and 4.3 kW at 295 K. To realize a pump inlet pressure of 295 K, the cavity leakage flow would have to be cooled by the water cooling system. We have also estimated the mass flow rates and pumping powers at the maximum up stream pressure of 13,842 kPa, but at elevated temperatures that vary from 325 K to 360 K. These results are illustrated in Figure 5-18. As just shown, there is considerable variation in the estimated leakage flow that varies from a low of 0.032 kg/s to a high value of 0.88 kg/s. The pumping power depends greatly on upstream pressure conditions, but also on the pump inlet temperatures. All of these values were unpredictable early in the program when the design decisions were made, hence the design was based largely on bounding estimates using the tools and correlations just described. This just points out the need to measure these leakage results, which was some of the very first test that was performed on the compression test-loop. In summary, the leak flow rates for this small scale proof-of-principle S-CO2 test loop are on the order of 1-2% of the total compressor mass flow rate per seal. However, because the pumping powers are relatively large (12 kW – 4 kW) there is a strong need to include features that reduce the pumping power such reducing the leakage flow rate by using better seals, or by pre-cooling the leakage flow prior to compressing it back into the loop. However, because of large uncertainties in the operating conditions the actual leakage rates need to be measured. It is also worth pointing out that this compression loop has only one compressor and one labyrinth seal. In later loops, the loops will have up to two compressors and two turbines with each wheel having a seal. Thus, measuring and understanding these loss mechanisms and developing good rules to scale these results to larger systems is important. Fortunately, based on our current understanding of the leakage flow for these designs, we expect that the fractional pumping power for large commercial systems to be much smaller than for this proof-of-principle test-loop, largely because more conventional sealing technologies can be used. Another reason is that in a larger system the generated power will grow as the radius squared, while the leakage flow rate grows proportional to the radius, thus it will be much easier to keep the fractional windage losses low in these larger systems. 66

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