Dynamic Simulation of a 800 MW el Hard Coal One-Through Supercritical Power Plant to Fulfill the Great Britain Grid Code

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1 Dynamic Simulation of a 8 MW el Hard Coal One-Through Supercritical Power Plant to Fulfill the Great Britain Grid Code HENNING ZINDLER E.ON Kraftwerke GmbH. Tresckowstaße 5 D-3457 Hannover GERMANY HEIMO WALTER Institute for Thermodynamics and Energy Conversion Vienna University of Technology Getreidemarkt 9, A-16 Vienna AUSTRIA ANDREAS HAUSCHKE and REINHARD LEITHNER Institute for Heat- and Fuel Technology Franz-Liszt-Straße 35 D-3816 Braunschweig GERMANY Abstract: - The dynamic behavior of a supercritical one-through hard coal power plant at fast load changes, which are required by the Great Britain Grid Code, was simulated with the aid of the dynamic simulation program Enbipro. The fast load changes will be achieved with the help of the primary measures turbine valve throttling and/or condensate stop. These primary measures are necessary during the time period which are required to increase significantly the coal combustion in the furnace. In the present article the results to the dynamic retention capacity of the boiler related to the thermal storage capacity of the steam mass and the steel mass are presented. The investigation has shown that 6% performance improvement can be obtained by the analyzed primary measures. Key-Words: - Great Britain Grid Code, Dynamic simulation, Turbine valve throttling, Condensate stop, One-through operation, Power plant, Enbipro 1 Introduction and problem formulation The Great Britain Grid Code [1] (GBGC) must be considered in the United Kingdom at the erection of new power plants and in particular at the erection of new supercritical coal fired power plants. A linear change of 1% of the power output up to 8% load is the demand of the GBGC as a reaction of the boiler in case of a frequency drop. This increase in active power output must be released increasingly within 1 seconds with a regeneration time of 2 minutes if the power plant works under part load condition between 55 and 8% of full load. The requirement diminishes linearly between 8 and 1% load. These extreme requirements are a result of the Great Britain transmission system, because the Great Britain power line works under "isolated operation" conditions. That means, that compared to the mainland of Europe a relatively low number of market participants exits in Great Britain and the individual consumption behavior of this participants is relatively unpredictable. The load change velocity of coal fired power plants is slow, therefore not all kind of measures of the power plant can be used to fulfill the above mentioned conditions of the GBGC. Consequently the techniques will be subdivided into primary and secondary measures. For the primary measures the short-term storage behavior of the power plant must be used. To this purpose the accumulator steam of the boiler by throttling, the steel mass of the boiler and the feed water tank (condensate stop) are counted among. These storages can be discharged by opening the throttled turbine valve or by condensate stop. These short-term or primary measures will be needed for the time lag which is given by the boiler to increase the evaporation as a secondary measure based on a increased firing power. During the regeneration cycle the furnace must be over-amplified because the accumulator capacity of the ISSN: ISBN:

2 power plant must be reloaded. The following questions are given for the accumulator capacity of the boiler: 1. How large is the steam mass accumulator capacity of the boiler at different load conditions and turbine valve positions? 2. How large is the influence of the thermal capacity of the steel mass? 3. Which time period after a frequency drop is necessary to throttle the turbine valve again? 4. Which time period after a frequency drop is necessary to get the boiler thermal stable? 5. How fast must be the rate of pressure change at the boiler outlet (opening of the turbine valve) and is thereby an effect on the lifetime of the boiler given? 6. Which influence to the power generating process is given by a higher injection rate of 2-3% related to the live steam mass? It should be noted at this place that a frequency drop normally will be advised by the network operator and the power plant will be operated in the annual mean at approximately 8% of full load. In the present article only the influence of the accumulator capacity of the boiler related to the steam mass and the thermal storage capacity of the steel mass will be analyzed. For the numerical calculation of the system response of the boiler by changing the power set point the dynamic simulation program Enbipro (Energy balancing program), which was developed at the Technical University of Braunschweig, Institute for Heat- and Fuel Technology, was used. 2 Model of the analyzed boiler Figure 1 shows the diagram of connections of the simulated supercritical hard coal fired one-through boiler (Benson type) with an electrical net capacity of approximately 8 MW el. The superheated steam mass flow at full load has an operation pressure of 285 bars and a live steam temperature of 6 C. The reheater outlet temperature is 62 C. The overall boiler volume of the working medium between the economizer and the superheater 3 (SH3) of the high pressure system is m 3. The high pressure system (HP) of the model consists of an economizer (ECO), an evaporator (EV), three superheater (SH1 to SH3) and a high pressure steam turbine. The intermediate pressure system (IP) consists of two reheater (RH1 and RH2) and the intermediate pressure turbine. Between the SH1 and SH2, SH2 and SH3 as well as RH1 and RH2 a spray valve is located. The intermediate (IP) and low pressure (LP) turbine are summarized in the used model to one turbine. Therefore only the IP reheat is included in the analyzed configuration of the boiler. The feed water mass flow is controlled by a predetermined curve, which is in this special investigation case a user defined input data for the used software tool Enbipro. In the present model the simulation of the parallel heating surfaces is neglect. That means that the simultaneously consideration of the convective heating surfaces, the furnace wall and the supporting tubes, is not included in the presented study. The furnace of the boiler is modeled as a convective heating surface with an adiabatic combustion temperature at the flue gas inlet. The geometry of the heating surfaces and the tube material mass of the boiler are included in the model. The headers and the connection tubes between the headers are neglect. The controller for the feed water as well as for the injection mass flow, which are not included in Fig. 1, are replaced by a control system. The control system for the spray valves was adjusted in such a way that in the base design the injected mass flow at both HP spray valves I1 and I2 was during the simulation 2% of the live steam and 1% at the IP spray valve I3. The turbine valve arranged upstream of the HP turbine is controlled with the help of a limited PI-controller (see Fig. 1). Fig. 1: Diagram of connections of the analyzed hard coal fired one-through boiler 2.1 Initial and boundary conditions for the simulation The dynamic behavior of a boiler during the start-up or a load change depends on the time difference between the shutdown and restart of the boiler as well as on the load change velocity. For the start-up from the cold condition (cold start, the system pressure is equal to the atmospheric pressure) or part load condition (warm start ISSN: ISBN:

3 or heavy load change) the admissible rate of temperature (pressure) change is mostly conditioned by the thermal stresses in the main steam headers and the material temperatures and temperature differences of the water walls. The present study was done for a heavy load change of a supercritical hard coal fired one-through boiler after a frequency drop by.5 Hz. The following boundary conditions are used for the dynamic simulations: Part load condition at a power ratio of P/P =.8. The condenser pressure is constant during the simulation at.3 bars. The feed water inlet temperature was constant at 3 C and the feed water inlet mass flow increases with a PT2 delay corresponding to the live steam mass flow rate. The flue gas temperature at the evaporator inlet is constant and identically to the adiabatic combustion temperature. The flue gas mass flow increases linearly with a time delay of 35 s. As initial condition for the dynamic simulations the result of a steady state calculation under part load condition (power ratio of P/P =.8) at full operation pressure was used. This results in a correct allocation of the vectors for the fluid properties and velocities for the flue gas as well as for the working medium. 3 Analyzed Parameters In the present paper the following parameters are analyzed to study its influence on the power output of the boiler: influence of the pressure difference between the turbine valve inlet and outlet Δp th on the mass of the working fluid accumulated inside the tubes of the boiler at different part load conditions. This investigation was done under steady state conditions (test case 1). influence of a linear load change of 6% within 1 s after simulation start at an pressure difference between the turbine valve inlet and outlet at simulation start of Δp th = 5 bars. The decrease of the boiler pressure by opening the turbine valve in front of the HP turbine is controlled by an IP controller. The upper limit of the pressure reduction velocity is 1 bar/s. (test case 2) influence on the power output of the boiler by a 1% higher injected mass flow rate at both HP spray valves I1 and I2 compared to the base configuration (test case 3). influence of a parameter combination of the test cases 2 and 3 on the power output of the boiler (test case 4). The investigation was done under a slightly less pressure difference between the turbine valve inlet and outlet at simulation start of Δp th = 4 bars compared to test case 2 with Δp th = 5 bars. The injected mass flow was the same as used in test case 3. The numerical simulation of the test cases 2 to 4 are done under part load condition at an power ratio of P/P =.8 at simulation start. As described above the GBGC requires a linear load change of 1% within 1 s. In the test cases 2 and 4 only a linear load change of 6% within 1 s are simulated. This is caused by the circumstance that at least 4 to 5% of the required additional power output should be produced with other primary measures, for example the condensate stop. A condensate stop means that no condensate flows from the condenser to the feed water tank and also no extraction steam will be taken from the turbines during the time period which is necessary for the power increase after the frequency drop. In the present study only the IP extraction tubes are closed. It must be mentioned at this point that the stress in the thick-walled structural components was not within the scope of the present analysis. The higher stress is a result of the fast pressure change which causes a change of the fluid and structural components surface temperature so that the temperature gradient inside this structural components will be increase. 4 Mathematical model The mathematical model [2] for the working medium is one-dimensional in flow direction and uses for the boiling region a homogeneous equilibrium model for the two-phase flow (based on the supercritical operation pressure of the boiler it was not necessary to use the two phase flow model in the current study). For a straight tube with constant cross section the governing equations in flow direction can be written for the conservation of the mass: ρ ρw + = (1) t x and for the conservation of the momentum: ρw ρww p p + = ρg x +. (2) t x x x Friction The density ρ and the velocity w are averaged values over the cross section of the tube. Considering the fluid flow in steam boilers, the thermal energy is much higher than the kinetic and the potential energy as well as the expansion work. Therefore, the balance equation for the thermal energy can be simplified to: ISSN: ISBN:

4 ρh ρhw U + = q& (3) t x A The heat exchange between fluid and wall is governed by Newton's law and the heat transfer through the wall is assumed to be in radial direction only. The heat transfer model used in Enbipro for the working medium is described in detail in [2]. The discretization of the partial differential equations for the conservation laws was done with the aid of the finite-volume-method. The pressure-velocity coupling and overall solution procedure are based on the SIMPLER [3] algorithm. To prevent checkerboard pressure fields a staggered grid is employed and for the convective term the UPWIND scheme is used. All other ordinary differential and algebraic equations, which are also used additionally in Enbipro, are solved with the help of a special method of the predictor-correction algorithm, the so called DASSL algorithm [4]. The DASSL algorithm belongs to a group of numerical initial value problem solving algorithms for implicit systems with the index zero and one in the form of r r r f ( t, y, y ) = (4) r r y( t ) = y (5) r r y ( t ) = y (6) and they are based on a backward differential equation method. To solve the differential algebraic system of equations the idea of Gear [5] was used which substitutes of the derivatives in the system of equations by approximated differences. The resulting algebraic system of equations can now be solved, for e. g. the time step t n+1, with the help of the Newton-method. The simplest variant is the implicit Euler-approach of a backward difference r r r r y n+ 1 yn f t 1, 1, n+ yn+ =. (7) tn+ 1 tn DASSL represents an extension of this idea. Instead of a linear approximation the backward differential equation method of the order k will be chosen, whereas k can receive the order between one and seven. The DASSL algorithm uses furthermore a control routine as a result of the solution behaviour for the variable increment and the variable order of the predictor polynomial on the basis of the fixed leading coefficient according to [6]. This control routine finds the balance between the calculating effort and the integration stability. For further information see [2], [8]. The implementation of the finite volume algorithm with the above described variant of the predictor corrector method will be done in Enbipro with the help of the adjoint-method, which is described in detail in [2], [7] and [8]. The calculation model for the turbines at full load is based on the polytropic change of state and is described in detail in [2]. For part load operation of the turbines the turbine inlet pressure is calculated with the help of Stodola's law [9], [1] ξ & ξ ξ p, ξ in T p in out, (8) m p 1 1 out = pin m& pin Tin,, pin with ξ + 1 κ = 2 η 1, (9) pol,t ξ κ which relates the flow rate through the turbine to the vapor conditions at the turbine inlet and outlet. 5 Discussion of the simulation results 5.1 Results of the steady state calculations to analyze the fluid accumulation inside the tube network of the boiler (test case 1) Fig. 2: Density distribution of the working fluid of the HP system at different operation pressures under steady state conditions Figure 2 shows the density distribution of the working fluid of the high pressure system outlined over the tube length at different part load conditions and valve positions of the turbine valve. The different graphs represent the density distribution under steady state conditions. The calculations are done for part load conditions of the boiler with an power ratio of P/P =.8 and.6 and a pressure difference over the turbine valve of Δp th = 5 and 2 bars (a smaller pressure difference indicates that the throttle valve is more open). During these calculations the condenser pressure was hold constant. The pressure at the turbine inlet was calculated using Stodola's law (8). It can be seen, that with a decreasing pressure difference between the inlet and outlet of the turbine valve the fluid density decreases. This is a result of the ISSN: ISBN:

5 lower operation pressure of the boiler upstream of the turbine valve at Δp th = 2 bars compared to the pressure difference of Δp th = 5 bars. Percent of full load power Pressure difference Δp th Mass of the working fluid in the boiler tubes 8 % 5 bar t 8 % 2 bar t 6 % 5 bar t 6 % 2 bar t Mass difference Δm 2.98 t t Table 1: Stored mass of the working medium as a function of the boiler operation conditions The required additional power (represented by the doted and dashed line in Fig. 3) will be achieved with a timedelay, because the turbine valve controller works in the this test case without a rate action. It can be seen also in Fig. 3 that the total power output is supplied by one third from the HP turbine and two third by the IP turbine. The unequal and time-delayed increase of the live steam mass flow in front of the HP and IP turbine results in different contributions of both turbines (see Fig. 3) to increase the power output of the boiler during the first 1 s after the frequency drop. Therefore the HP turbine must be overshoot the power output during this time period to compensate the time-delay of the IP. This can be seen in Fig. 4 and 5 by a higher live steam mass flow and pressure in front of the HP turbine and also in Fig. 3. Table 1 shows for the different analyzed operation conditions the determined mass of the working medium inside the tubes of the boiler as well as the mass difference of the working medium at a constant ratio P/P and different pressure differences between the turbine valve inlet and outlet. A comparison of the mass difference shows that with reduction of the boiler power output Δm increases. 5.2 Results of the dynamic boiler simulation with throttling of the mass flow at the inlet of the HP turbine (test case 2) Fig. 3: Power output of the turbines at load change Figure 3 shows the time evolution of the power output of the plant for the HP and IP turbine after a 6% linear load change. The load change was finished 1 s after simulation start. The calculation was done under the part load condition of P/P =.8 and a pressure difference over the turbine valve at simulation start of Δp th = 5 bars. Fig. 4: Mass flow of the working medium and flue gas at different points of the boiler The system response of the mass flow of the flue gas and the working medium after a 6% linear load change is presented in Fig. 4. As a result of the load change the live steam mass flow in front of the HP turbine increases and achieves its peak value approximately 15 s after simulation start. The overshooting of the steam mass flow results also in a higher power output of the HP turbine over the same period of time. During the time difference between 15 and 3 s, the superheated mass flow at the HP turbine inlet decreases. After this period the mass flow aspirates to the steady state condition after the load change. The mass flow at the IP turbine inlet shows a slightly deviant behavior compared to the HP turbine inlet, because the IP turbine is arranged downstream of the HP turbine and the two reheaters RH1 and RH2. Therefore the higher mass flow through the IP turbine will increase with a time-delay. This can be seen in Fig. 4 by the slower increase of the mass flow in front of the IP turbine. Fig. 5 shows the pressure evolution over the time in selected points of the analyzed boiler. The curve for the ISSN: ISBN:

6 pressure of the working medium at the inlet of the IP turbine is scaled with a constant factor of 5. The scaling is necessary for a better presentation of the time-delay of the pressure increase. Fig. 6: Temperature of the working medium at the HP and IP turbine inlet and flue gas temperature at boiler outlet Fig. 5: Pressure of the working medium at different points of the boiler By using the stored mass of the working medium inside the tubes of the boiler to increase the power output for a short time period the throttled turbine valve must be opened. The opening of the HP turbine valve results in 1. a lower operation pressure of the boiler and 2. in an increase of the pressure in front of the HP turbine. This higher pressure is a result of the increasing steam mass flow (see Fig. 4) through the turbine valve as consequence of Stodola's law, which can be seen in Fig. 5. The overshooting of the HP inlet mass flow (see Fig. 4) is represented in Fig. 5 by the overshooting of the HP inlet pressure and an accelerated decrease of the feed water and SH3 pressure. In the present test case the pressure change velocity by opening the turbine valve is 1 bar/s. This is about two times faster than the reference value given by [11]. Approximately 1 s after simulation start the turbine valve is complete opened and no stored mass inside the tubes of the boiler is available for further primary measures. This can be seen also in Fig. 5 by the difference between the SH3 outlet and the HP turbine inlet pressure. At this time the pressure difference between turbine valve inlet and outlet at full opened turbine valve is equivalent to the pressure difference between the graphs of the SH3 outlet and the HP turbine inlet. With achieving the full opened turbine valve the mass flow through the turbine decreases and consequently also the power output of the turbine and in last consequence of the power plant decreases too until the secondary measures are sufficiently high enough. A temperature change of the working medium is linked with the change of the operation pressure of the boiler. Based on the increasing pressure at the inlet of the HP turbine the temperature of the working medium also increases (see Fig. 6). The change of the IP pressure during the load change is small compared to that of the HP system. Therefore the temperature at the IP turbine inlet is approximately constant. Based on this circumstance the thermal storage behaviour of the boiler is not used in a sufficient way. 5.3 Results of the dynamic boiler simulation by a higher injection mass flow (test case 3) Fig. 7: Power output of the turbines at a higher injected mass flow rate at both HP spray valves I1 and I2 As presented in test case 2 the thermal storage behaviour of the boiler is not used in a sufficient way as a primary method to increase the power output of the steam generator. Therefore the potential of opening the spray ISSN: ISBN:

7 valve on the power output of the boiler will be analysed in this paragraph. The additional injected mass flow rate at both HP spray valves I1 and I2 was during the simulation approximately 1% higher ( 1 kg/s) compared to the base design. During the simulation of this test case all controller were deactivated. Figure 7 presents the power output of the HP and IP turbine as well as the total power output of the steam generator plant. The total power output of the plant is 15 s after simulation start approximately 1.3% higher compared to the value at the beginning of the simulation. The temperature at the inlet of the turbine valve, which can be seen in Fig. 8, decreases during this time period from 594 C to 585 C as a result of the mass injection. spray valves I1 and I2 compared to the base configuration). Fig. 9: Power output of the turbines at load change The power output of the HP and IP turbine as well as the total power output of the power plant is presented in Fig. 9. A doted and dashed line with the required 6% higher power output is also included in Fig. 9. It can be seen that with the same performance but lower throttling the required improvement in the power output of the plant can also be achieved. Fig. 8: Temperature of the working medium at the HP and IP turbine inlet and flue gas temperature at boiler outlet at test case 3 This lower medium temperature leads also to a lower surface temperature of the tubes, which can result to a lower lifetime of the power plant. The fast change of the fluid temperature can have also an influence on the operation of the steam turbine which is not known at this moment (additional thermal stress in the turbines). 5.4 Results of the dynamic boiler simulation with throttling of the mass flow at the inlet of the HP turbine and a higher injection mass flow (test case 4) In the current investigated case a combination of the studied parameters of the test cases 2 and 3 on the power output of the boiler was analysed. This investigation was done under a slightly less pressure difference between the turbine valve inlet and outlet at simulation start of Δp th = 4 bars compared to test case 2 with Δp th = 5 bars. The injected mass flow was the same as used in case 3 (1% higher injected mass flow rate at both HP Fig. 1: Pressure of the working medium at different points of the boiler (test case 4) Approximately 9 s after simulation start the thermal storage of the boiler is empty, which can be seen in Fig. 1 by the pressure difference between inlet of the HP turbine and the outlet of SH3. At this time the turbine valve is completely opened and the pressure difference between the turbine valve inlet and outlet is equivalent to the pressure difference between the SH3 outlet and the HP turbine inlet ( 6 bars). The simulation result shows that a frequency drop can be handled by the supercritical hard coal fired onethrough boiler by using the thermal storage capacity of ISSN: ISBN:

8 the boiler. The method presented in test case 4 has compared to test case 2 additional reserves available as a result of the lower pressure difference between the turbine valve inlet and outlet at simulation start. 6 Conclusion At the erection of new power plants in the United Kingdom the GBGC must be considered. The GBGC regulates that a linear change of 1% of the power output up to 8% load must be done as a reaction of the boiler in case of a frequency drop. In the present article the influence of the accumulator capacity of a once through super critical boiler related to the steam mass and the thermal storage capacity of the steel mass on the power output of the boiler after a frequency drop was analyzed. The investigation has shown that 6% performance improvement can be obtained by the analyzed primary measures. The remaining 4% additional power output, which are necessary to fulfill the GBGC must be produced with other primary measures, for example the condensate stop. In the next future the numerical results should be verified by measured data at a real power plant. 7 Nomenclature A Cross section area [m 2 ] f r function vector [-] g x Component of the gravity in direction of the tube axis [m/s 2 ] h Spec. enthalpy [J/kg] k Order of the backward differential equation method [-] m& Mass flow [kg/s] m& Mass flow at full load [kg/s] Δm Difference of stored mass [t] p Pressure at part load [Pa] p in Pressure at turbine inlet [Pa] p in, Pressure at turbine inlet at full load [Pa] p out Pressure at turbine outlet [Pa] p out, Pressure at turbine outlet at full load [Pa] p Pressure at full load [Pa] Δp Pressure difference [Pa] Δp th Pressure difference between throttle inlet and outlet [Pa] P Power at part load [MW] P Power at full load [MW] q& Heat flux [W/m 2 ] t Time [s] T in Fluid temperature at turbine inlet [K] T in, Fluid temperature at turbine inlet at full load [K] U Perimeter [m] w Fluid velocity [m/s] x Length [m] y r Vector for differential or algebraic variables [-] r y Derivation of the vector for the differential or algebraic variables [-] y r Vector of the initial values for the differential or algebraic variables [-] r y Derivation vector of the initial for the differential or algebraic variables [-] ρ Density [kg/m 3 ] η pol,t Polytropic efficiency of the turbine [-] κ Polytropic exponent [-] References: [1] National Grid, CONNECTION CONDITIONS, idcode/gridcodedocs/, May 28 [2] Zindler, H., Dynamic power plant simulation - coupling of the finite volume algorithm and the predictor-corrector method with the adjoint-method, Progress-report VDI, VDI-Publishing Company, Düsseldorf, 28 (in press, in German). [3] Patankar, S. V., Numerical Heat Transfer and Fluid Flow, Series in Computational Methods in Mechanics and Thermal Sciences, Hemisphere Publ. Corp., Washington, New York, London 198. [4] Brenan, K. E. Campbell, S. L and Petzold, L. R., Numerical Solution of Initial-Value Problems in Differential-Algebraic Equations. SIAM Classics Series, Elsevier Science Publishing Co. 2 nd edition, 1996 [5] Gear, C. W, Numerical Initial Value Problems in Ordinary Differential Equations. Prentice Hall, Inc. Englewood Cliffs, New Jersey, Stanford, California 1971 [6] Jackson, K. R. and Sacks-Davis, R., An alternative implementation of variable step-size multistep formulas of stiff ODE's, ACM Trans. Math. Software, Vol. 6, 198 [7] Martins, J. R. R. A., Alonso, J. J., and Reuther, J., Aero-Structural Wing Design Optimization Using High-Fidelity Sensitivity Analysis; CEAS Conference on Multidisciplinary, Aircraft Design and Optimization, Cologne, Germany, June 25-26, 21 [8] Martins, J. R. R. A., Alonso, J. J., and Reuther, J., A Coupled Adjoint Sensitivity Analysis Method for High-Fidelity Aero-Structural Design, Optimization and Engineering, Vol. 6, No. 1, 25, pp [9] Pfleiderer C. and Petermann, H., Strömungsmaschinen, Springer Verlag, 24 ISSN: ISBN:

9 [1] Hauschke, A., Dynamic simulation and optimization of a heat recovery steam generator, Master thesis, Technical University of Braunschweig, 27 (in German). [11] Strauss, K., Kraftwerkstechnik: Zur Nutzung fossiler, nuklearer und regenerativer Energiequellen, Springer Verlag, 26 ISSN: ISBN: