Course of Energy Conversion A 2014/2015

Transcription

1 Course of Energy Conversion A 2014/2015 Prof. G. Valenti 3 rd project Diverse Heat recovery steam cycles Index: 1. Introduction; 2. 1-level combined cycle analysis 3. 2-levels combined cycle analysis 4. 3-levels with reheating combined cycle analysis

2 1. Introduction The aim of the project is to compare, with respect to the previous project, different possible configurations of the heat recovery steam generator in a combined cycle plant, at fixed gas turbines conditions. The plant is in fact optimized to work under the constraints of mass flow rate, power generated (and hence the first law efficiency) and exhaust outlet temperature of the gas turbine. Therefore, for the gas turbine we will take those assumptions, with respect to the constraints (highlighted in green): Simple Cycle parameters inlet mass flow rate 649,35 kg/s compression ratio 20 - TIT 1299 C T exhausts 546,7 C Net power 235,65 MW plant efficiency 0,338 - All the values listed here will be constant through the analysis of the diverse plant configurations. Notice that the mass flow rate is computed in order to grant the same power of the plant and the same exhaust mass flow rate to the heat recovery steam generator of the project 2. The quantities considered have been adjusted in order to have the most similar plant configuration as possible.

3 2. 1-level combined cycle analysis The first case we will take in account is the configuration that follows: # Despite the presence of the condensate removal line in the scheme, we will not consider this operation (so the mass flow rate will be considered equal to zero). The plant is in the exact same configuration as the one treated in the previous project, hence we will take the evaporation pressure at the steam drum as a constraint, in order to maximize the net electrical power output of the steam cycle. The Heat Recovery Steam Generator (which, for sake of simplicity, will be called HRSG from now on) so, consist of a single line with the pressure set at 32,6 bar at the steam drum. The other constraints we are considering are: condensing pressure: 0,05 bar; deaerator pressure: 2 bar; pressure losses through the economizer (Δp/p): 10 %; ΔT pp (pinch-point): 10 C; ΔT ap (approach): 25 C; ΔT sc (subcooling): 10 C; And as before, this values hold for all the calculation we will make during this report; notice that the only values we are changing, are the pressures with respect to the different configuration. With the help of the software Turbogas, we find the following values regarding the performance of the analysed plant:

4 Combined cycle - 1L Evaporation pressure 32, bar Steam cycle efficiency - η sc 0, Thermal recovery η th,rec 0, Recovery - η rec 0, Second law efficiency - η II 0, Steam cycle net power 95,17 MW Exhaust temperature at the stack 143,2 C Δη HRSG 0, Δη stack rejection 0, W rev available of the exhaust gases 169,715 MW The steam cycle net power is very close to the one computed in the previous project: this matching is quite encouraging as it means that the optimization was successful; The second law efficiency is definitely good, but we expect it to be even higher as the levels of pressure raise; The efficiency losses are computed with respect to the reversible power of the exhaust gases, which is by the way constant through the calculations (as we will not modify the gas turbine and the pressure losses through the HRSG the temperature at the stack will not modify); The thermal recovery is computed as the ratio between the recovery and the steam cycle efficiency : η th,rec = η rec η steam cycle The temperature at the stack is quite high due to the single-level plant, which does not allow us to recover in a proper way the heat from the exhausts flow (as we can see from the value of the thermal recovery): we can notice that easily from the T-Q diagram.

5 3. 2-levels combined cycle analysis The second configuration we will inspect is the one that provides two levels of pressure, respectively: Low pressure level: 6.3 bar; High pressure level: 76 bar; We expect an improvement in the recovery efficiency (and so in the thermal recovery) due to an higher power production (but a lower steam cycle efficiency, as we know that the pressure of the first case is optimized to grant the highest power production). The HRSG we are now considering has the following scheme: We will consider all the plant parameters fixed but the pressures at the steam drums. Once again, notice we will not take in account any condensate removal from the steam turbine, meaning a lower efficiency. Let us now perform some calculations: Combined cycle - 2L Evaporation pressure - HP 76 bar Evaporation pressure - LP 6,3 bar Steam cycle efficiency - η sc 0, Thermal recovery η th,rec 0, Recovery - η rec 0, Second law efficiency - η II 0, Steam cycle net power 105,05 MW Exhaust temperature at the stack 93,8 C Δη HRSG 0, Δη stack rejection 0, W rev available of the exhaust gases 169,715 MW

6 As we expected, there is an increase in the recovery efficiency, and so an increase of the net power production; dividing the heat absorption in two stages affects directly the losses in the HRSG, that diminish slightly (from 18% to 13%). Because of this better recovery, we notice that the temperature at the stack is lowered of a huge amount (near to fifty degrees) and that is both good and bad. If from one side this means a better recovery, on the other we must be aware that at this temperature the dangerous species contained in the flue gases (as NOx, and less SOx as we are considering a gas turbine) can condensate forming acids with water dissolved in the air. Hence, we suggest taking particular care in designing the gas turbine with respect to the abatement of emissions; as an example, we should consider the employment of a dry-low-nox combustor and the SCR at the stack (which, unfortunately would cost a bit in terms of pressure losses).

7 4. 3-levels w/ RH combined cycle analysis The last configuration we are taking in account is absolutely the most complicate, expensive yet more efficient possible; let us remember that enhancing again the number of levels would be non-sense from a practical point of view, as the benefits would not be justified by the enormous increase of technological complication and of the cost of the plant. The scheme is: Even in this case, no condensate removal considered and all fixed parameters but the pressures. The pressures considered are: Low pressure level: 3 bar; Intermediate pressure level: 25 bar; High pressure level: 150 bar; The reheating is performed at the intermediate pressure level. To clarify the reading of the scheme, we will report here the main thermodynamic properties of the cycle points. Point of the cycle m' [kg/s] T [ C] P [bar] H [kj/kg] S 1 94,08 32,09 0, ,4 3, ,08 110, ,88 4, ,81 120, ,89 10,643 4 not considered 0 232,5 2, ,77 11, ,08 120, ,48 5, ,68 123, ,35 5,081

8 7-133, ,75 5, ,68 133, ,52 10, ,68 208, ,25 10, ,68 327, ,15 11, , ,29 11, ,25 133, ,28 5, ,25 213, ,26 5, , ,21 6, ,25 223, ,24 9, ,25 327, ,81 10, ,67 525, ,84 10, ,14 133, ,85 5, ,14 223, ,73 6, ,14 332, ,99 7, , ,17 7, ,14 342, ,15 8, ,14 525, ,2 9, ,42 299, ,9 10, ,08 32,9 0, ,15 11,296 Notice the difference in the mass flow rates going from the low-pressure level to the high-pressure one: there is a big increase due to the inflow of the intermediate/high pressure streams. # This is crucial when considering, for example, the repowering of an old coal-based power plant: while here we have increasing flow rate, in the old plant there would be decreasing mass flow rate due to bleedings (for the preheating line);the steam turbines are so designed for decreasing mass (and volumetric) flow rates. Hence, in that case we must be aware in confirming that the reduced flow rates of the steam turbine are respected. Performing the calculations, we obtain: Combined cycle - 3L +RH Evaporation pressure - HP 150 bar Evaporation pressure - IP 25 bar Evaporation pressure - LP 3 bar Steam cycle efficiency - η sc 0, Thermal recovery η th,rec 0, Recovery - η rec 0, Second law efficiency - η II 0, Steam cycle net power 113,26 MW Exhaust temperature at the stack 87,6 C Δη HRSG 0, Δη stack rejection 0, W rev available of the exhaust gases 169,715 MW

9 In this configuration, the all the efficiencies increase greatly: the reheating process is highly beneficient; this configuration is in fact considered as the state of the art, but it is not usually applied but in the large plants as it is extremely expensive and technologically complicated. As we saw, the efficiencies are improved, the power generated is greater as well as the power recovered from the gas flow. The losses over the HRSG and at the stack diminish slightly: we notice a very low temperature at the chimney (87.6 C); the same consideration we spoke about before hold true even in this case. Let us have a look more specifically at the second law losses in efficiency: Entropy analysis - 3L + RH Δη,cycle inflow heat (HRSG) 0, Δη,steam turbine 0, Δη,dissipation waste, condenser 0, Δη,exh.gases at the chimney 0, Δη,th,mech,electr,aux 0, Tot: 0, second law efficiency: 0, We can see easily that with this configuration, the losses at the chimney are definitely low if compared to the others, confirming the high quality of the recovery process (counterbalanced by higher losses of the inflow heat transfer to our cycle, unfortunately). The overall analysis, however, shows a big improvement in the cycle if compared to the other two configurations, making it (at least under a thermodynamical point of view) the most suitable one.

10 COMPARISON 1L 2L 3L + RH Water flow rate 87,53 100,89 94 kg/s Steam cycle efficiency 0,3228 0, , Thermal recovery 0, ,8532 0, Recovery 0,2469 0,2725 0, Second law efficiency 0,5608 0,6195 0, Steam cycle net power 95,17 105,05 113,26 MW Exhaust temperature at the stack 143,2 93,8 87,6 C Deta HRSG 0,1751 0, , Deta gas stack rejection 0, , , Wrev available (gases) 169, , ,715 MW