Maritime University of Szczecin
Akademia Morska w Szczecinie
2011, 27(99) z. 1 pp. 5–11 2011, 27(99) z. 1 s. 5–11
The influence of chosen operation parameters on the efficiency
of supercritical power plant
Wpływ wybranych parametrów pracy na efektywność bloku
nadkrytycznego
Łukasz Bartela, Janusz Kotowicz, Sebastian Michalski
Silesian University of Technology, Faculty of Mechanical Engineering Institute of Power Engineering and Turbomachinery
Politechnika Śląska, Wydział Mechaniczny Technologiczny Instytut Maszyn i Urządzeń Energetycznych
44-100 Gliwice, ul. Konarskiego18
e-mail: lukasz.bartela@polsl.pl, janusz.kotowicz@polsl.pl, sebastian.michalski@polsl.pl
Key words: steam cycle, thermodynamic analysis, operation parameters Abstract
In the paper the evaluation of applied solutions, as well as the operational parameters in context of achieved efficiency of steam cycle of a supercritical power plant was carried out. The analysis for the referred structure of a power plant was made. The thermodynamic computational model was created in GateCycle software. In the calculations, the basic assumption was the maintenance of constant gross power of power plant. For this purpose, in the model the flow of live steam entering steam cycle was changed. In the first step the assumptions for three cases of unit were determined. First of all, the variants differ from each other with live and reheated steam parameters. The parameters for the two examined cases can be regarded as possible to be currently achieved in a modern supercritical power plants, while the third case represents a perspective variant. Based on relevant data implemented into the computational model, the relevant results for quantities characterizing the operation of examined units were generated. Among others, in the analysis the values of following quantities were determined: steam cycle efficiency, gross overall efficiency and specific consumption of heat. In the paper the assumptions, as well as the results are shown in the form of tables. The conclusions from comparison of results for individual solutions are also formulated.
Słowa kluczowe: obieg parowy, analiza termodynamiczna, parametry pracy Abstrakt
W artykule dokonano oceny stosowanych rozwiązań, jak również parametrów pracy, w kontekście uzyskiwa-nej sprawności obiegu parowego bloku nadkrytycznego. Analiza wykonana została dla określouzyskiwa-nej struktury elektrowni. Termodynamiczny model obliczeniowy utworzono w programie GateCycle. W trakcie obliczeń podstawowym założeniem było utrzymanie mocy elektrycznej brutto bloku na stałym poziomie. W tym celu w modelu uzmienniano strumień pary świeżej wprowadzanej do części parowej układu. W pierwszej kolejno-ści określono założenia liczbowe dla trzech wariantów bloku. Warianty przede wszystkim różnią się między sobą parametrami pary świeżej oraz wtórnie przegrzanej. Parametry dla dwóch badanych przypadków mogą być traktowane, jako aktualnie możliwe do osiągnięcia w nowoczesnych blokach na parametry nadkrytyczne, z kolei wariant trzeci reprezentuje rozwiązanie perspektywiczne. Na podstawie danych wprowadzonych do modelu obliczeniowego wygenerowano odpowiednie rezultaty dla wielkości charakteryzujących pracę bada-nych układów. W analizie określono wartość między innymi dla takich wielkości, jak: sprawność obiegu pa-rowego, sprawność wytwarzania energii elektrycznej brutto oraz jednostkowe zużycie ciepła. Zarówno dane, jak i wyniki przedstawione są w pracy w formie tabel. W artykule sformułowano wnioski wynikające z ze-stawienia wyników uzyskanych dla poszczególnych rozwiązań.
Introduction
Currently observed in the world trend in efforts to reduce emissions, especially of greenhouse gas-es, contributs to significant changes in the direction of the development of energy technologies [1]. This is extremely important in the field of coal technolo-gies, due to importance of these fuels in the energy balances of many countries, including Poland, as well as to significant emission of CO2 per unit
elec-tricity production (about 2.5 times greater than in the area of natural gas-fired combined cycle). In the area of coal technology there are two very impor-tant research directions aiming to bring the reduc-tion of CO2 unit emission, thus, in consequences, to
the reduction of global emission:
search of low-energy consuming carbon capture technologies (including searching of new solu-tions, optimization of known solusolu-tions, also in the area of integration of the CCS unit with power plant);
increasing the efficiency of electricity genera-tion, including optimization of power plant, both in the area of its structure, as well as in area of operation parameters [2].
Among carbon capture technologies three direc-tions are developed:
pre-combustion technology, post-combustion technology, oxy-combustion technology.
In the area of coal technologies all of solutions can be used. However, the first solution is predis-posed for IGCC systems [3], in which there is the possibility of generating carbon dioxide before combustion of synthesis gas. The fuel before enter-ing the combustion chamber is subjected to carbon sequestration. Due to lower gas stream from which the carbon dioxide is removed such a separation process is connected with less energy consumption. The next technology is based on removing carbon dioxide from flue gases leaving the power system. The post-combustion technology is predisposed for the conventional coal-fired power plants. In the area of post-combustion technology, as in the case of pre-combustion, the research on absorption and adsorption techniques, as well as membrane and cryogenic separation are realized [4]. In the area of clean coal technology large hopes are associated with oxy-combustion technology, of which the principal purpose is combustion of coal in an oxy-gen-rich atmosphere in order to eliminate from the exhaust gases the inert gas (nitrogen). In this case the exhaust gases leaving the steam boiler consists mainly of carbon dioxide and steam, so the carbon
capture process is much less energy intensive. In research area of oxy-combustion technology, currently the solutions aiming for decreasing ener-gy consumption connected with oxygen production in air separation unit are searched for [5, 6, 7].
The results presented in the paper were realized within the framework of the Strategic Project “Advanced Technologies for Energy Generation: Oxy-combustion technology for PC and FBC boi-lers with CO2 capture”. In the paper the results of
analysis of the steam cycles of energy generation units which are the object of study in these Project are shown. These steam cycles will be the basis for creation of the models of the whole oxy- -combustion power plants. In the paper the results of analysis including the influence of different solu-tions of steam cycles, and thus their assumed para-meters, on the energy effectiveness evaluation indi-cators are shown.
Description of the model of steam cycle and the assumptions for analysis
In the paper the following variants of steam cycles, which particularly can be characterized by giving their gross power, as well as the parameters of live and reheated steam, are analyzed:
W1 variant: gross power: 460 MW, steam para-meters: 600/600°C, 29/4.8 MPa;
W2 variant: gross power: 600 MW, steam para-meters: 600/620°C, 29/5 MPa;
W3 variant: gross power: 600 MW, steam para-meters: 650/670°C, 30/6 MPa.
In analysis the W3 variant due to high parame-ters of live and reheated steam can be treated as perspective solution. Such high parameters are not currently used due to the limited strength of the materials used in the production process of high temperature heat exchangers.
For analysis, two structures of steam cycle were assumed. The main solution here is to use feed wa-ter pump driven by an electric motor. In the figure 1 the solution with feed water turbopump by using dashed line was marked. Each of the variants described above for two structures was analyzed.
The steam cycle consists of a boiler with rehea-ter (K), steam turbine (consisting of high-pressure part WP, intermediate-pressure part SP and low-pressure part NP), electric generator (G), condenser (KND) (for structure with turbopump the additional condenser KND2), deaerator (ODG), condensate pump (PS), feed water pump (or turbopump) (PWZ), seven regenerative heat exchangers (four low-pressure (NP) and three high-pressure (WP))
and steam cooler (SCH). In the case with turbo-pump the deaerator and the turboturbo-pump are fed from the same bleeding of intermediate-pressure part of steam turbine. The low-pressure regenerative heat exchangers are fed from the blades of intermediate- and low-pressure part, while the high-pressure re-generative heat exchangers are feed from the blades of high- and intermediate-pressure part of the steam turbine.
The values of isentropic efficiencies, which are shown in table 1, depend on the enthalpy at the inlet to the turbine (iin), in the blade or at the outlet from
the turbine (iout) and enthalpy of steam after
isen-tropic expansion in blade or at the outlet from the turbine (iouts). The relations between these quantities
is described by the following equation:
s out in out in i i i i i (1)
The gross powers, which are shown in table 1, depend on the internal power of steam turbine (Ni WP, Ni SP, Ni NP), mechanical losses in steam
tur-bine (Nm) and losses in electric generator
deter-mined based on assumed generator efficiency (g):
i i i m
gel N N N N
N TP WP SP NP (2)
The gross powers for each part of steam turbine are determined based on the balance equations writ-ten for group of stages located between individual inlets, blades or outlets.
The assumptions relating to the operational parameters for three variants are shown in table 1.
The assumptions shown in table 1 are based on current trends in the construction of modern coal-fired supercritical power plants. The assumptions are based on the published literature, in particular [8]. The assumptions were used for calculations performed on the model of steam cycle built in the GateCycle software. Here, for various components of the steam cycle the predisposed models were used. Among the assumed quantities listed in table 1 only the gross power (Eq. (2)) is the quantity which determines the scale of power plants. On the basis of this quantity in the model the steam flow is determined. For calculation of this flow in algo-rithm the special implemented macro is used. Here, the flow calculated for next iterative loop depends on the current flow of live steam (
m1 i), the as-sumed gross power ((Nel,b)zał) and the gross powerdetermined for current iterative loop ((Nel,b)i):
i el,b el,b i i i N N m , m , m zał 1 1 1 1 05 05 (3)Fig. 1. Diagram of steam cycle of a power plant Rys. 1. Schemat obiegu pary wodnej elektrowni
In the case of structure with turbopump, the additional turbine is fed from the same steam bleed-ing of steam turbine as deaerator. The steam flow required to feed this turbine is determined by the energy consumption of feed water pump and the losses connected with transfer of energy from tur-bine to a pump.
The results of computations
The calculations presented in this section of the article were carried out with the use of the model of steam cycle build in GateCycleTM program. The
calculations were performed for the assumption concerning three considered systems (indicated as
Table 1. Characteristics quantities for three analyzed variants Tabela 1. Charakterystyka ilościowa trzech analizowanych wariantów
Quantities Sym- bols Values Unit
W1 W2 W3
Gross power Nel,B 460 600 600 MW
Temperature of live steam leaving the boiler t1 604.9 604.9 654.9 °C
Pressure of live steam leaving the boiler p1 30.1 30.1 31.1 MPa
Temperature of live steam at inlet to the steam turbine t2 600 600 650 °C
Pressure of live steam at inlet to the steam turbine p2 29 29 30 MPa
Temperature of reheated steam leaving the boiler t4 602.4 622.4 672.4 °C
Pressure of reheated steam at the inlet to the steam turbine p5 4.8 5 6 MPa
Temperature of reheated steam at the inlet to the steam turbine t5 600 620 670 °C
Deaerator operating pressure p24 1.2 MPa
Condenser operating pressure p19 0.005 MPa
Pressure at the outlet of the condensate pump p31 1.6 MPa
Temperature of feedwater t42 297 297 310 °C
Isentropic efficiency of group of stages of high-pressure steam turbine ηi WP 90 %
Isentropic efficiency of group of stages of intermediate-pressure steam turbine ηi SP 93 %
Isentropic efficiency of group of stages of low-pressure steam turbine ηi NP 86 %
Isentropic efficiency of last group of stages ηi NP1 81 %
Generator efficiency ηG 99 %
Steam turbine mechanical losses ΔNmTP 4.6 6.0 6.0 MW
Isentropic efficiency of pumps ηi P 85 %
Efficiency of regenerative heat exchangers ηW 99.5 %
Efficiency of steam cooler ηSCH 99.5 %
Efficiency of deaerator ηODG 99.5 %
Relative pressure drops in steam pipelines to regenerative heat exchangers, steam cooler and
deaerator ζPI 2.0 %
Relative pressure drop in steam pipeline from steam cooler to WP1 regenerative heat exchanger ζ59–60 1.0 %
Relative pressure drop of water flowing through the low-pressure regenerative heat exchangers ζNP 6.0 %
Relative pressure drop of water flowing through the high-pressure regenerative heat exchangers
and steam cooler ζWP 0.5 %
Relative pressure drop of working medium in boiler ζ42–01 11.0 %
Relative pressure drop of steam in steam reheater ζ03–04 3.0 %
Relative pressure drop in reheated steam pipelines ζ12–03
ζ04–05
1.4
2.4 %
Relative pressure drop between SP part and NP part of steam turbine ζ06–07 0.5 %
Temperature differences in low-pressure and high-pressure regenerative heat exchangers ΔTsNPi
ΔTsWPi
3
3 K
Difference between temperature of condensate at outlet and temperature of water at inlet in low- and high-pressure regenerative heat exchangers
ΔTaNPi,
ΔTaWPi
10
10 K
Increase of temperature of water in low-pressure regenerative heat exchangers ΔTNPi 30 K
Increase of temperature of water in WP1 regenerative heat exchanger ΔTWP1 30 30 40 K
Increase of temperature of water in WP2 regenerative heat exchanger ΔTWP2 37.04 39.60 41.16 K
Increase of temperature of water in WP3 regenerative heat exchanger ΔTWP3 30.57 28.00 16.22 K
W1, W2, W3). For each of the system two struc-tures were considered, the structure with the use of the feedwater pump powered by electric motor and the structure with the use of the feedwater pump powered by steam turbine. In order to thermody-namically evaluate the respective variants, the fol-lowing quantities were used: efficiency of steam cycle, gross overall efficiency, gross power, speci-fic consumption of heat and specispeci-fic consumption of chemical energy of fuel. The gross overall effi-ciency and the specific consumption of chemical energy of fuel was determined assuming that the steam cycles work in cooperation with a steam boiler with the efficiency of ηk = 94%. For the
determination of the listed evaluation indicators during the calculations, following parameters were determined: heat supplied to the steam cycle, heat discharged in the condenser, heat loss in the rege-nerative heat exchangers and deaerator, heat loss in the live and the reheated steam pipelines and total amount of heat discharged from the steam cycle. The dependencies for the majority of these quanti-ties are shown below.
The efficiency of steam cycle depends on the heat supplied to the steam cycle (Qd) and the heat discharged from the steam cycle (Qw). The depen-dence between these quantities is expressed by the following formula: d w d ob Q Q Q η (4)
The gross overall efficiency depends on the gross power of a power plant (Nel,b) and the chem-ical energy stream of fuel (Q ). The formula for ch
calculations of these efficiency is as follows:
ch el,b el,b Q N η (5)
The flow of the heat supplied to the steam cycle depends on the flow of the live steam (m ), the 1
flow of the reheated steam (m ), the enthalpy of 4
the live steam (i1), the enthalpy of the feed water
entering the boiler (i42), the enthalpy of steam
enter-ing the reheater (i3), and the enthalpy of the
re-heated steam (i4). The relationship between these
quantities is expressed by the following expression:
1 42
4
4 3
1 i i m i i
m
Qd (6)
The flow of the heat discharged from the steam cycle depends on the heat discharged in condenser (QKND and eventually QKND2), the heat loss in the regenerative heat exchangers and steam cooler (QPRWSCH), the heat loss in the deaerator (QsODG) and the heat loss in the live and the reheated steam pipelines (QsR). The relationship between these quantities is expressed by the following formula:
ODG SCH PRW R KND s s w Q Q Q Q Q (7) The specific consumption of heat depends on the gross power of power plant (Nel,b) and the heat
sup-plied to the steam cycle (Qd).The formula for the calculation of this indicator is as follows:
el,b d
c N
Q
q 3600 (8)
The specific consumption of chemical energy of fuel depends on the gross power of power plant (Nel,b) and the chemical energy stream of fuel (Q ). ch
Table 2. Selected characteristic quantities for three system variants for steam cycle with pump driven by an electric motor
Tabela 2. Charakterystyka ilościowa wybranych wielkości trzech wariantów systemu obiegu pary wodnej z pompy napędzanej silni-kiem elektrycznym
Quantity Symbol Values Unit
W1 W2 W3
Efficiency of steam cycle ob 49.78 49.97 51.30 %
Gross overall efficiency el,b 47.44 47.60 48.84 %
Gross power el,b 460.00 600.00 600.0 MW
Specific consumption of heat qq 7133.48 7109.59 6929.1 kJ/kWh
Specific consumption of chemical energy of fuel qch 7588.81 7563.40 7371.4 kJ/kWh
Heat supplied to the steam cycle Qd 911500.8 1184932.2 1154850.8 kW Heat discharged in the condenser QKND 453121.4 586825.7 556195.0 kW Heat loss in the regenerative heat exchangers and deaerator QWYM 1589.0 2044.3 2069.2 kW Heat loss in the live and the reheated steam pipelines Q R 3026.0 3905.2 4168.4 kW
It can be calculated based on the following formula: k c el,b ch en_ch η q N Q q 3600 (9)
The values calculated with the use of equations (4)–(9) and others aiming to compare the efficiency of work of the considered structures of the steam cycle for three considered systems are presented in tables 2 and 3.
Conclusions
After the results of calculations presented in tables 2 and 3 were analyzed, the following conclu-sions were made:
Replacement of pump by turbopump in structure of steam cycle caused:
an increase of the efficiency of steam cycle by about 0.27 percentage points and of the overall gross efficiency by about 0.02 percentage points for all variants;
a decrease of the specific consumption of heat, specific consumption of chemical energy of fuel, heat supplied to the steam cycle and heat dis-charged from the steam cycle.
Increase of the temperature of the reheated steam from 600°C to 620°C and of the pressure of the reheated steam from 4.8 MPa to 5.0 MPa for steam cycle caused:
an increase of the efficiency of steam cycle by 0.19 percentage points regardless of the used structure;
an increase of the overall gross efficiency by 0.16 percentage points for the structure with pump and by 0.15 for the structure with turbo-pump;
a decrease of the specific consumption of heat by 23.89 kJ/kWh for the structure with pump and by 23.59 kJ/kWh for the structure with tur-bopump;
a decrease of the specific consumption of chem-ical energy of fuel by 25.41 kJ/kWh for the structure with pump and by 25.10 kJ/kW for the structure with turbopump.
Increase of the temperature of the live steam from 600°C to 650°C, of pressure of the live steam from 29 MPa to 30 MPa, of the temperature of the reheated steam from 620°C to 670°C and of pressure of the reheated steam from 5.0 MPa to 6.0 MPa for steam cycle caused:
an increase of the efficiency of steam cycle by 1.33 percentage points for the structure with pump and by 1.32 percentage points for the structure with turbopump;
an increase of the overall gross efficiency by 1.24 percentage points regardless of used struc-ture;
a decrease of the specific consumption of heat by 180.49 kJ/kWh for the structure with pump and by 180.07 kJ/kWh for the structure with tur-bopump;
a decrease of the specific consumption of chem-ical energy of fuel by 192.00 kJ/kWh for the structure with pump and by 191.56 kJ/kWh for the structure with turbopump;
a decrease of amount of the heat supplied to the steam cycle by 30,081.40 kW for the structure with pump and by 30,011.60 kW for the struc-ture with turbopump;
a decrease of amount of the heat discharged from the steam cycle by 30,342.50 kW for the
Table 3. Selected characteristic quantities for three system variants for steam cycle with turbopump
Tabela 3. Charakterystyka ilościowa wybranych wielkości trzech wariantów systemu obiegu pary wodnej z turbopompy
Quantity Symbol Values Unit
W1 W2 W3
Efficiency of steam cycle ob 50.06 50.25 51.57 %
Gross overall efficiency el,b 47.47 47.62 48.86 %
Gross power el,b 460.00 600.00 600.00 MW
Specific consumption of heat qq 7129.46 7105.87 6925.80 kJ/kWh
Specific consumption of chemical energy of fuel qch 7584.53 7559.43 7367.87 kJ/kWh
Heat supplied to the steam cycle Qd 910986.3 1184311.0 1154299.4 kW Heat discharged in the condenser QKND 450304.3 583280.4 552797.4 kW
Heat loss in the regenerative heat exchangers and deaerator QWYM 1588.2 2043.2 2068.2 kW Heat loss in the live and the reheated steam pipelines Q R 3024.3 3903.1 4166.4 kW
structure with pump and by 30,194.70 kW for the structure with turbopump.
Results of study show, that considering only the parameters of steam, achieving of 50 percent of steam cycle efficiency, can be reached for both considered structures. However, taking into account the assumed boiler efficiency, achieving a level of 50% gross overall efficiency of power plant is not possible. The literature indicates many additional solutions that allow to obtain increase of efficiency. These include the use within the range of low temperature system of waste heat, coal drying, optimizing in heat recovery direction of the power plant structure, or the use of the double reheating of steam. These guidelines will be considered in the next steps of research.
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Acknowledgements
Scientific work was supported by the National Centre for Research and Development, as Strategic Project PS/E/2/66420/10 “Advanced Technologies for Energy Generation: Oxy-combustion
technolo-gy for PC and FBC boilers with CO2 capture”. The
support is gratefully acknowledged.
Recenzent: dr hab. inż. Andrzej Adamkiewicz, prof. AM Akademia Morska w Szczecinie
