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In Figure 9 a very clear relationship is presented, where, with the increase in LMTD in recuperator, the percent of waste heat transferred to CO2 also increases. This results from the fact that with larger values of LMTD in recuperator, more energy is “wasted” and “taken” from the system at the cooler. To compensate this, a correspondingly greater amount of waste heat from the flue gases must be supplied to the system. Energies 2020, 13, 2447 13 of 18 In Figure 10 the relationship between the LMTD and waste heat utilization rate is presented. 3.8 3.4 3.0 Waste heat utilisation rate (%) ΔT=5K 2.6 ΔT=7K ΔT=10K ΔT=20K 2.2 0 20 40 60 80 100 120 LMTD in recuperator (K) Figure 10. Relationship between the LMTD and waste heat utilization rate for varied temperature Figure 10. Relationship between the LMTD and waste heat utilization rate for varied temperature differences between CO2 and the exhaust gases at the inlet and outlet of the heat exchanger. differences between CO2 and the exhaust gases at the inlet and outlet of the heat exchanger. Waste heat utilization rate is an energy efficiency indicator, which combines CE and PWHT. The flow so that the given amount of waste heat could be transferred to the S-CO2 system. In the pressure value of WHUR increases both with the growth in CE and PWHT. The data presented in Figure 10 and temperature range assumed for the system operation, the CO2 enthalpy value (for a constant shows that the WHUR value increases slightly with the decrease of LMTD in recuperator. temperature) decreases with increasing pressure. In the case of the discussed system, this is especially ΔT=3K The CO2 mass flow was not the assumed value. CO2 mass flow was calculated as the required Results presented in Figures 8–10, show that for the adopted operating conditions, the increase noticeable after passing CO2 through a recuperator (cold side). Assuming a constant heat flux to be in internal efficiency has a greater impact on the WHUR than decrease in percent of waste heat transferred to the system, the smaller enthalpy at the main heat exchanger inlet causes that the CO2 transferred to CO2 (PWHT). The increase of WHUR is much more significant when the temperature velocity in the exchanger must be lower, and hence also the mass stream. When it comes to the increase difference between CO2 and flue gases in the heat exchanger decreases. This is due to the fact that in the mass stream of CO2 with the growth of exhaust gases temperature at the heat exchanger outlet, both CE and PWHT grow with the temperature difference drop. By overlapping these growths, the this is due to the fact that less energy from the outside is required for transfer to the system (the merit increase in WHUR is evident. of the recuperator, which raises the temperature of CO2 at the heat exchanger inlet). The less energy from the outside that must be transferred to the system, the higher the CO2 velocity in the exchanger 5. Results and Discussion that can be achieved, and hence the mass stream. Figures 4–7 show that if the temperature of exhaust As mentioned in the introduction, no other publications were found that presented an analysis gases at heat exchanger outlet is higher, the cycle efficiency is better. of S-CO2 cycle application for the waste heat recovery at a natural gas compressor station. Detailed 4c.o2m. IpnavreistoignatoiofncoynclTeemefpfiecriaetnucrye DiniffdeirceantcoersbevtwaleuenesthbeeHtwoteaendiCtsoladpSpildiecaotfiothnesRiencuvpaeriaotuors industries is pointless, because these are different operating parameters (power of devices, waste heat To investigate the reasons for the cycle efficiency growth related to the increase of temperature temperature, etc.). For example, Kim et al. [6] studied the supercritical CO2 power cycle for landfill of exhaust gases at heat exchanger outlet, it was decided to additionally examine the influence of gas fired gas turbine bottoming cycle. They concluded that for the simple recuperated cycle the net temperature difference between the hot and cold sides of the recuperator. For this purpose, the authors produced work is 2.18 MWe with 29.98% cycle net efficiency. For comparison, in this paper maximum defined logarithmic mean temperature difference of the recuperator given by Equation (11): net power system was 47 kW with cycle efficiency 14%. However, it should be noticed that Kim et al. T′′ −T −T −T′′ assumed exhaust temperature 519.69 °C and mass flow rate of the flue gases of 21.3 kg/s. In this work 2362 it was, respectively, 343.23 °C and L3.M55TkDg/=s. Similarly, the maximum system operating pressure w(1a1s) (T′′ −T3) LN 5 ′′ taken as 27.6 MPa (22.5 MPa in this paper). Taking this into account, it was decided to compare the values of selected S-CO2 efficiency (T6−T2 ) indicators with ORC efficiency indicators, analyzed for the same gas compressor station. Kowalski et where: al. [2] analyzed the application of ORC system at Jarosław II natural gas compression station. The LMTD—logarithmic mean temperature difference of the recuperator (K), ′′ ORC system was proposed for the production of eclectic energy. They conclude that the average net T2 —temperature of CO2 at the recuperator cold side inlet (K), T3—temperature of CO2 at the recuperator cold side outlet (K), ′′ T5 —temperature of CO2 at the recuperator hot side inlet (K), T6—temperature of CO2 at the recuperator hot side outlet (K). Calculations of the energy efficiency indicators were carried out for different LMTD values, for a constant turbine inlet pressure of 22.5 MPa. In addition, the same calculations were repeated forPDF Image | Supercritical CO2-Brayton Cycle Nat Gas Compression Station
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