Combined Cycle Start-Up Procedures: Dynamic Simulation and ...

COMBINED-CYCLE START-UP PROCEDURES: DYNAMIC SIMULATION AND MEASUREMENT

Nicolas J. Mertens1, Falah Alobaid1, Bernd Epple1, Hyun-Gee Kim2

1Department of Energy Systems and Technology (EST), Technische Universität Darmstadt, Germany 2Thermal & Fluid Research Team, Doosan Heavy Industries & Construction Co. Ltd., Korea

ABSTRACT

The daily operation of combined-cycle power plants is increasingly characterized by frequent start-up and shutdown procedures. In addition to the basic requirement of high efficiency at design load, plant operators therefore acknowledge the relevance of enhanced flexibility in operation – in particular, fast start-ups – for plant competitiveness under changing market conditions. The load ramps during start-up procedure are typically limited by thermal stresses in the heat recovery steam generator (HRSG) due to thick-walled components in the high pressure circuit. Whereas conventional HRSG design is largely based on simple steady-state models, detailed modelling and dynamic simulation of the relevant systems are necessary in order to optimize HRSG design with respect to fast start-up capability. This study investigates the capability of a comprehensive process simulation model to accurately predict the dynamic response of a triple-pressure heat recovery steam generator with reheater from warm and hot initial conditions to the start-up procedure of a heavy-duty gas turbine. The commercial combined-cycle power plant (350 MWel) was modelled with the thermal-hydraulic code Apros. Development of the plant model is based on geometry data, system descriptions and heat transfer calculations established in the original HRSG design. The numerical model is validated with two independent sets of measurement data recorded at the real power plant, showing good agreement.

INTRODUCTION

Combined cycle power plants (CCPP) have received much recognition in the last decades for high efficiency, fast load response and comparatively small environmental impact [1]. In the combined-cycle process, the waste heat of a gas turbine (GT) unit is absorbed by a heat recovery steam generator (HRSG) installed downstream in the flue gas path. The steam is used in a Rankine bottoming cycle, which generates additional power in the steam turbine (ST). A triple-pressure subcritical HRSG system with reheater, where GT and ST units are combined in a 1+1 configuration, is considered state of the art [2]. Nominal efficiency of modern CCPPs amounts to more than 60 % and is expected to further increase in the future due to research on innovative GT cooling concepts and high-temperature resistant materials [3]. Fast response capability of conventional power plants is a prerequisite for increasing shares of renewable feed-in [4] and thus represents a competitive advantage. Accurate calculation of the transient system behavior is integral part of the CCPP design process, with particular regard to control design [5]. Due to the inherently small time scale of GT dynamics relative to the bottoming cycle (BC) [6], the key endeavor in CCPP transient performance evaluation is the estimation of the BC dynamic response to the changes in GT exhaust flow and temperature [7]. In essence, this is the subject of the present work. The relevant systems of the reference plant are defined and the modelling

Proceedings of the ASME 2016 Power Conference POWER2016

June 26-30, 2016, Charlotte, North Carolina

POWER2016-59286

1 Copyright © 2016 by ASME

process is explained. The model is subjected to GT start-ups from hot and warm initial conditions, which represent the dominant contributions to overall lifetime consumption for critical components. Validation of the simulation results is performed and deviation from the measured transients of the real plant is discussed.

REAL POWER PLANT

The combined-cycle power plant considered in this work consists of a gas turbine and a vertical heat recovery steam generator, which is installed downstream in the flue gas path. Figure 1 shows a simplified flow diagram of the power plant,

corresponding to the system boundaries of the eventual model. For the sake of clarity, several auxiliary systems such as control valves, attemperators and steam turbine bypasses are neglected. The HRSG is divided in three distinct pressure levels (low, intermediate and high pressure) and the individual heat exchangers are arranged to match the temperature curves of flue gas and water-steam side. Part of the HP economizer and the IP economizer as well as IP superheater and LP superheater are arranged in parallel. Waste heat from the passing flue gas is absorbed by the serrated-fin tubes, generating live steam for the corresponding turbine sections. Nominal conditions of the thermodynamic process for flue gas and steam mass flows are summarized in Table 1.

Exhaust gas conditions HRSG outlet steam conditions Condenser Combined-cycle power

HP IP (after RH) LP

587.3kg/s 97.7 bar 23.4 bar 4.1 bar 83 mbar 350 MW

628 °C 567.1 °C 567.2 °C 293.4 °C 42.2 °C

80.2 °C 78.4 kg/s 83.4 kg/s 9.9 kg/s

HP EVAPORATOR

IP EVAPORATOR

LP EVAPORATOR

LP ECO

HP ECO 1

IP ECO // HP ECO 2

HP ECO 3 LP SH // IP SH

HP ECO 4

REHEATER 1

REHEATER 2

HP SH 1

HP SH 2

G

From gas turbine exhaust

To stack

LP drum

IP drum

HP drum

Condensate pump

Steam turbine

Condenser

LP circuit

IP/RH circuit

HP circuit

Table 1: Nominal process parameters of the power plant

Figure 1: Schematic flow diagram of the combined-cycle power plant


Main components of the water-steam side are: i. Forced circulation evaporator circuits HP, IP

and LP ii. Three steam turbine stages HP, IP and LP

iii. Reheater section after expansion in the HP turbine

iv. Condenser v. Condensate pump and boiler feed pumps HP,

IP Due to the fact that some steam is diverted after expansion in the HP turbine for fuel gas preheating and sealing steam, the increase of the steam mass flow from HP stage to the reheater is small.

MODELLING

The heat recovery steam generator is modelled using the dynamic process simulator Apros developed by VTT Finland [8]. Apros includes extended component libraries and numerical algorithms for solving the non-linear equation systems that describe the time-dependent operating state of the plant. The software tool was chosen on account of the highly detailed modelling of water-steam processes due to its origin from nuclear safety analysis.

Flow Model

The HRSG model is built with thermal-hydraulic modules based on the heterogeneous (six-equation) flow model, complemented by control circuits. The heterogeneous model is particularly suitable for two-phase water-steam systems with notable slip between the phases. It is based on the governing mass, momentum and energy balances but weighted by the respective volume fractions α of the two separate phases. The

subscript k may either adopt the values l for liquid org for gas, yielding a total of six partial differential equations as follows.

Mass balance:

Momentum balance:

,

Energy balance:

, ,

,

The subscripts i and w refer to phase interface and wall

boundary, respectively. F and Q denote friction force per volume and heat flow. The sum on the right-hand side of the momentum equation is a source term accounting for a particular component, such as the steam turbine. The pressure drop in the turbine is described by Stodola’s cone law.

, ,

, ,

The subscripts 1 and 2 refer to the nodes before and after the turbine module, and 0 is the turbine design point. The flow resistance of the turbine is calculated as function of the Stodola coefficient K. Similarly, in the energy balance the turbine is represented as heat sink of the downstream node.

Water/steam

Modelling

Flue gas

Heat structure

Heat transfer

Calculation branch

Calculation node

Figure 2: Discretization of counter-flow heat exchanger


Figure 3: Flue gas conditions and combined-cycle power

/ /

ξ, ηis and h0 denote resistance coefficient, isentropic efficiency, and reference enthalpy. The equation system is closed by empirical correlations that account for various exchange phenomena at the phase boundary such as interfacial friction and evaporation/condensation. For a more detailed explanation of the flow model, the reader is referred to Mertens et al. [9].

The equation system is solved numerically by discretization of the six partial differential equations with respect to space and time and linearization of non-linear terms, see Siikonen [10]. Discretization in time applies fully implicit scheme, whereas spatial discretization uses first-order upwind scheme as well as a staggered grid in order to prevent a non-physical pressure solution. Different correlations are inserted for closing the resulting linear equation system, which is solved for each control volume by iteration. At the beginning of the iteration cycle, material properties are computed as function of pressure and enthalpy. The algorithm advances to the next time step if the solution has converged, i.e. the residual mass error is negligibly small.

Power Plant Model

The model of the above-described plant is divided in flue gas side and water-steam side; the latter is further subdivided in the respective pressure circuits and the steam turbine section. At the inlet of the flue gas side, the gas turbine is simplified as ideal mass source described by the following process parameters: mass flow, temperature and exhaust gas composition. Mass flow and temperature are imposed as transient boundary conditions according to measurement data, and flue gas composition is assumed constant for all loads. After passing the complex arrangement of serrated-fin heat exchanger tubes, cold exhaust gas reaches the ambient-pressure boundary condition at the outlet.

Figure 4: HP steam parameters

The water-steam side is structured according to the three pressure circuits at low, intermediate and high pressure level, corresponding to 4 bar, 23 bar and 98 bar, and the turbine section (including condenser). For illustration of the model structure, a draft of the LP circuit is included in the Annex. A small part of the expanded HP steam is diverted for fuel gas preheating, which increases the efficiency of the overall combined-cycle process. The diverted steam mass flow is assumed to be a linear function of GT load. Condenser pressure (pcond = 83 mbar) is the only physical boundary condition of the water-steam side. The water-steam components were modelled with real geometry data so that meaningful estimation of thermal stresses in critical parts (in particular HP drum, HP and RH headers) is possible. All heat exchangers except the evaporators are counter-flow type. Figure 2 illustrates the modelling of a counter-flow heat exchanger with an equivalent one-dimensional module: Flue gas path and water-steam tubes are discretized in equally-spaced control volumes with one calculation node in the center and a calculation branch between two adjacent nodes. In the same way, the tube walls are discretized as a radial structure of heat nodes and heat branches, which represent the steel mass and the thermal conductivity of the tube material. Convective heat transfer from flue gas flow to serrated fins on the one side and from inner tube surface to water-steam flow on the other side are calculated by heat transfer modules. Tube mass is increased in order to account for external fins and gas-side heat transfer is adjusted following the ESCOA revised correlation [11] for serrated-fin tubes in staggered arrangement. For each pressure circuit, three-element PI feedwater control is used in order to adjust the position of the feedwater control valve and prevent both dry-out of the drum during shutdown

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Figure 5: HP drum pressure and feedwater flow

and water expulsion to the superheater during start-up. The positions of the ST admission valves are imposed as function of time and the positions of the ST bypass valves are adjusted by drum pressure control. The live steam temperatures are controlled by HP and RH spray attemperators before admission to the steam turbine. A more exhaustive description of the model can be found in Mertens et al. [12].

START-UP PROCEDURES

Start-up procedures are roughly classified as hot starts after overnight shutdown, warm starts after weekend shutdown and cold starts after a shutdown of several days. However, these categories are too broad for practical purposes and operating transients are determined as a function of initial “cold” metal temperatures of the thick-walled components [13]. The start-up in terms of steam admission is conducted as follows: Both the main steam control valve and the bypass control valve are closed at GT ignition, causing fast pressure build-up in the live steam system. First steam generation is accommodated by ST bypass and condenser, while the steam admission valve remains closed. When a preset GT holding point is reached, the bypass valve starts opening in order to maintain a constant pressure. Once sufficient steam quality is confirmed, the main steam valve is opened while the position of the bypass valve is changed in reverse. Start-up procedure is completed when the turbine bypass is fully closed, which means that the entire steam mass flow is routed to the turbine. This point corresponds to approximately 90 % of nominal load for a start-up to full load, so that CCPP load may still increase after completion of the actual start-up sequence. Simulation of the start-up procedures is performed by imposing flue gas mass flow, GT exhaust gas temperature and composition as boundary conditions on the virtual plant model, and calculating the system response of the bottoming cycle. The response of steam mass flow, temperature and pressure as well as feedwater mass flow and drum pressure is presented in the

Figure 6: RH steam parameters

following. The analysis is focused on the high-pressure circuit and reheater since these sections are subjected to the highest temperature gradients. Complemented by the data for intermediate-pressure circuit and low-pressure circuit outlined in the Annex, these results allow comprehensive validation of the power plant model with two independent sets of measurement data. In all figures, solid lines indicate measurement and dashed lines indicate simulation.

Hot Start-up

Figure 3 shows the GT boundary conditions for hot start-up, measured CCPP load and calculated steam turbine power. The considered period in time can be divided in pre-start phase (minutes 0-15), start-up (minutes 15-86), load-following operation (minutes 86-245) and design operation (minutes 245-360). As safety precaution, pre-start the gas turbine is driven by the generator at approximately 20 % of nominal rotation speed to blow a small air mass flow through the HRSG and purge the flue gas system of any residual hydrocarbons. The start-up is initiated with GT ignition, after which the turbine rapidly accelerates to nominal rotation speed in approximately five minutes. Synchronization is completed at minute 45 and the loading process is started. At minute 50, sufficient steam quality is reached and first HP steam is admitted to the steam turbine. Ramp-up to the desired load is continued after the bypass stations are fully closed. Between minutes 86 and 245, the combined-cycle plant is operated in load-following operation at high part loads above 75 % before switching to design operation near full load (350 MWel).

In Figure 4, measured and calculated HP steam conditions are compared. During pre-start and start-up procedure the simulation results generally show qualitative agreement. However, noticeable quantitative deviations are observed, which will be explained in the following.

Time [min]

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Figure 7: Flue gas conditions and combined-cycle power

The remaining pressure in the live steam system after overnight standstill amounts to approximately 40 bar and initial temperature drops to 270 °C since the HP superheater is cooled by purging air during pre-start. Shortly after GT ignition, the simulation shows a temperature jump to 450 °C that precedes the measured temperature increase by an average of six minutes and is also characterized by a steeper transient (48 K/min compared to 17 K/min). The model underestimates the thermal capacitance of the heat recovery steam generator because auxiliary systems (such as safety vents, draining systems, steam sampling lines) and supporting structures in thermal contact with the primary water-steam circuits as well as ambient losses are not considered. Thus the tube material heats considerably faster and first steam generation in the HP evaporator loop is predicted six minutes earlier than observed in the real plant. Accordingly, the calculated pressure transient precedes the measured pressure increase. In load following operation, the calculated steam mass flow is in quantitative agreement with measurement but the steam temperature shows a discrepancy due to attemperator control. Whereas the simulated HP water injection keeps live steam temperature almost constant at the ideal set point of 567 °C, for use of the real injection component a minimum mass flow is required. Thus the temperature control tends to overshoot when engaged, causing negative temperature peaks during load changes. Steam pressure shows qualitative agreement between t = 85 min and t = 245 min, where the transients are well described by simulation but the time discrepancy is passed on from start-up phase. All HP steam conditions show good quantitative agreement (maximum deviations: -4.7 %, 1.3 % and -0.8 % for steam flow, pressure and temperature) when the plant operates at nominal load.

Figure 5 shows the comparison for HP drum pressure and HP feedwater mass flow. Since the drum inventory is always in

Figure 8: HP steam parameters

saturation state, the drum temperature is a simple function of pressure and therefore not presented. The calculated drum pressure shows good quantitative agreement for design operation and qualitative agreement in pre-start and start-up phase. Before first steam is admitted to the steam turbine at t = 50 min, bypass control aims to keep the drum pressure at approximately 40 bar. However, the detailed operating characteristics of the turbine bypass valves were not available in this study so that generic valves with a standard driving time (tvalve = 10 s) were used. This modelling uncertainty is reflected by the pressure deviations after GT ignition, when the initial steam generation is routed through the HP bypass valve. The simulation of feedwater mass flow agrees well with measurement, particularly for load-following operation and at design load. In contrast, the mass flow during start-up process is characterized by massive flow oscillations with maximum amplitude of 72 kg/s that could not be predicted by the calculation. When first steam is generated, water is discharged to the drum due to the decrease of average density in the HP evaporator tubes. The mass flow controller attempts to compensate the resulting level jump instantly and responds by closing and opening the feedwater control valve in rapid succession. Mass flow control in the model is less rigid so that a limited deviation from level set point is tolerated initially and then reduced over time, resulting in a smooth flow characteristic and reduced control action of the valve.

In Figure 6, the comparison of simulation results and measurement data for reheater temperature and pressure is presented. The steam mass flow is omitted in the figure because the corresponding measuring point is not available in the real plant. However, 90 % of the reheated steam flow are

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Figure 9: HP drum pressure and feedwater flow

constituted by the HP steam mass flow (with all bypass stations closed) so that similar behavior can be expected. The calculated pressure of the reheater section shows qualitative agreement for start-up and load-following phases and good quantitative agreement for design operation (maximum deviation: -4.1 %). Reheater temperature at GT ignition is higher compared to the HP superheater since the superheater is directly cooled by flue gas at the end of the previous shutdown and by purging air in the pre-start phase. By analogy with the superheater, the calculated reheat temperature at start-up precedes the measured temperature by an average of six minutes due to the under prediction of thermal capacitance in both reheater and HP superheater. This results in early steam production and correspondingly early increase of system pressure. The temperature shows very good quantitative agreement in design operation (maximum deviation: -0.96 %). In simulation the ideal live steam temperature (567 °C) is set by the RH attemperator, while plant temperature control tends to overshoot due to the minimum mass flow required by the injection nozzle component.

Warm Start-up

Figure 7 shows the boundary conditions of GT exhaust mass flow and exhaust temperature for warm start-up as well as the CCPP load profile. The considered period in time can be divided in pre-start phase (minutes 0-15), start-up (minutes 15-140) and operation at high loads (minutes 140-726). The start-up sequence is initiated with GT ignition. Synchronization is completed at minute 30 and the loading process is started. At minute 61, sufficient steam quality is reached and first HP steam is admitted to the steam turbine. Ramp-up to the desired load is continued after the bypass stations are fully closed. Between minutes 140 and 726, the combined-cycle plant is operated at high part loads above 75 % except for the initial loading process from start-up.

Figure 10: RH steam parameters

In Figure 8, measured and calculated HP steam conditions are compared. During pre-start and start-up procedure the simulation results generally show qualitative agreement. The initial conditions in the live steam system are approximately 5 bar and 155 °C. GT ignition yields a temperature jump to 470 °C in simulation that precedes the measured temperature offset by six minutes. Due to the steeper transient (70 K/min compared to 20 K/min), this time lag increases to a maximum of 30 minutes. The model underestimates the thermal capacitance of the heat recovery steam generator since auxiliary systems and supporting structures as well as ambient losses are neglected. Accordingly, first steam generation is predicted eight minutes earlier and the calculated pressure transient precedes the measured pressure increase. In high-load operation (>75 % part load), all HP steam conditions are in quantitative agreement (maximum deviations: -9.5 %, 3.8 % and 5.3 % for steam flow, pressure and temperature) with measurement. Deviation of the steam mass flow is highest when the measured temperature differs from the set point temperature, which suggests that in the real plant the HP attemperator is engaged. In particular, the pressure transient of the steam system is accurately predicted.

Figure 9 shows the validation for HP drum pressure and HP feedwater mass flow. With the exception of the initial start-up ramp, the predicted drum pressure is in quantitative agreement with measurement. Similar to hot start-up, larger deviation is observed during bypass switchover due to unavailable information on the physical valve characteristics. The simulation of feedwater mass flow agrees well with measurement, particularly for operation in the high-load range. However, the observed mass flow oscillations during start-up are not reflected in the results due to less rigid setting of the mass flow controller.

In Figure 10, the comparison for reheater temperature and pressure is presented. The calculated pressure of the reheater

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section shows qualitative agreement for start-up and quantitative agreement for high-load operation (maximum deviation: 8.5 %). By analogy with the HP superheater, the calculated reheat temperature at start-up precedes the measured temperature by an average of six minutes. This results in early steam production and correspondingly early increase of system pressure. In high loads > 75 %, the temperature is in good quantitative agreement with measurement (maximum deviation: 4.3 %).

CONCLUSION

Detailed simulation results for each pressure stage are presented and compared with measurement data of the power plant, yielding the following conclusions:

Calculated steam mass flows, pressures and temperatures are in good agreement with two separate sets of measurement data. The results confirm that the developed model can successfully reproduce the dynamic behavior of the real plant.

However, the thermal inertia of the water-steam side is underestimated since auxiliary systems and supporting structures in contact with the primary flow paths are not considered. This is reflected in the calculations by early prediction of first steam generation as well as higher temperature gradients during the start-up procedures.

This work highlights the importance of complete information on components and control circuits for modelling. In this work the settings of the actual feedwater valve controllers were not available, which explains that the new control design yields a smooth characteristic of the feedwater mass flow rather than the observed mass flow oscillations.

The calculated temperature gradients could e.g. be applied in the design phase of new combined-cycle plants as boundary conditions for FEM analysis of thermally stressed components [14], resulting in a more accurate (but still conservative) estimation of material fatigue. Hence, typical safety coefficients for these components could be reduced to decrease wall thickness and improve plant flexibility. The developed model will serve as reference for prospective dynamic studies on model reduction for start-up optimization and on the transient performance of supercritical HRSGs. These results show the potential of dynamic simulation for the further development of combined-cycle power plants in particular, as the plants are characterized by a high level of automation and standardization.

ACKNOWLEDGMENTS

Mr. Mertens gratefully acknowledges funding by Deutsche Forschungsgemeinschaft (DFG) within the framework of the Darmstadt Graduate School of Energy Science and Engineering (GSC 1070).

NOMENCLATURE

ck = two-phase friction multiplier [-] DH = hydraulic diameter [m] E = rate of entrainment [-] F = force/volume [N/m3] f = friction coefficient [-] gz = gravitational component in z-direction [m/s2] h = static enthalpy [kJ/kg] h0 = total enthalpy [kJ/kg] k = interfacial heat transfer coefficient [kg/s]

Q = heat flow/volume [kW/m3] p = pressure [bar] R = rate of stratification [-] t = time [s] u = fluid velocity [m/s] z = spatial coordinate [m] α = void fraction [-] Γ = mass exchange rate [kg/(m3 s)] ρ = density [kg/m3]

SUBSCRIPTS

comp = component-specific (valve, pump, orifice) el = electric g = gas phase i = index or phase interface ig = interaction between interface and gas phase ik = interaction between interface and liquid/gas phase il = interaction between interface and liquid phase k = liquid or gas phase l = liquid phase ns = non-stratified flow s = stratified flow sat = saturation sp = single phase wk = interaction between wall and liquid/gas phase


REFERENCES

[1] B. Lu, M. Shahidehpour, Short-term scheduling of combined cycle units, IEEE Transactions on Power Systems, 19 (2004) 1616-1625. [2] R. Kehlhofer, B. Rukes, F. Hannemann, F. Stirnimann, Combined-Cycle Gas \& Steam Turbine Power Plants, Pennwell Books, 2009. [3] S. Hada, M. Yuri, J. Masada, E. Ito, K. Tsukagoshi, Evolution and Future Trend of Large Frame Gas Turbines: A New 1600 Degree C, J Class Gas Turbine, in: ASME Turbo Expo 2012: Turbine Technical Conference and Exposition, American Society of Mechanical Engineers, 2012, pp. 599-606. [4] H. Zindler, H. Walter, A. Hauschke, R. Leithner, Dynamic Simulation of a 800 MWel Hard Coal One-Through Supercritical Power Plant to Fulfill the Great Britain Grid Code, in: 6th IASME/WSEAS International Conference on Heat Transfer, Thermal Engineering and Environment, Rhodes, Greece, 2008, pp. 184-192. [5] P.O. Larsson, F. Casella, F. Magnusson, J. Andersson, M. Diehl, J. Akesson, A framework for nonlinear model-predictive control using object-oriented modeling with a case study in power plant start-up, 2013 IEEE Conference on Computer Aided Control System Design (CACSD). Part of 2013 IEEE Multi-Conference on Systems and Control, (2013) 346-351. [6] J. Shin, Y. Jeon, D. Maeng, J. Kim, S. Ro, Analysis of the dynamic characteristics of a combined-cycle power plant, Energy, 27 (2002) 1085-1098. [7] S.C. Gülen, K. Kim, Gas Turbine Combined Cycle Dynamic Simulation: A Physics Based Simple Approach, Journal of Engineering for Gas Turbines and Power, 136 (2014) 011601-011601. [8] APROS Advanced Process Simulation Software. See also http://www.vtt.fi/?lang=en, in, 2016. [9] N. Mertens, F. Alobaid, R. Starkloff, B. Epple, H.-G. Kim, Comparative investigation of drum-type and once-through heat recovery steam generator during start-up, Applied Energy, 144 (2015) 250-260. [10] T. Siikonen, Numerical method for one-dimensional two-phase flow, Numerical Heat Transfer, Part A: Applications, 12 (1987) 1-18. [11] H. Walter, R. Hofmann, How can the heat transfer correlations for finned-tubes influence the numerical simulation of the dynamic behavior of a heat recovery steam generator?, Applied Thermal Engineering, 31 (2011) 405-417. [12] N. Mertens, F. Alobaid, T. Lanz, B. Epple, H.-G. Kim, Dynamic simulation of a triple-pressure combined-cycle plant: Hot start-up and shutdown, Fuel, 167 (2016) 135-148.

[13] C. Ruchti, H. Olia, K. Franitza, A. Ehrsam, W. Bauver, Combined Cycle Power Plants as ideal solution to balance grid fluctuations. Fast Start-up Capabilities, in: Proc. of 43th Colloquium of Power Plant Technology, 2011, pp. 18-19. [14] H. Hack, Z. Fan, A. Seltzer, J. Alvarez, Advanced Methods of HRSG Design for Life Cycle Optimization Under Fast Startups, POWERGEN International, (2012) 11-13.


ANNEX A: Model structure Figure 11: Schematic of the LP circuit

ANNEX B: Additional results, hot start-up Figure 12: IP steam parameters

Figure 13: IP drum pressure and feedwater flow

From condensate pump

LPECO 1LPECO 2LPECO 3LPECO 4LPECO 5LPECO 6

LPECO 7 LPECO 8 LPECO 9 LPECO 10 LPECO 11 LPECO 12

ECO recirculation

ECO bypass

LP CONTROL VALVE

LP drumLP SUPERHEATER

To LP turbine

To IP economizer

To HP economizer

LP EVAPORATOR 2

LP EVAPORATOR 1

LP EVAPORATOR 3

IP BOILER FEED PUMP

HP BOILER FEED PUMP

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Figure 14: LP steam parameters

Figure 15: LP drum pressure and feedwater flow

ANNEX C: Additional results, warm start-up

Figure 17: LP steam parameters

Figure 18: LP drum pressure and feedwater flow

Figure 19: IP drum pressure and feedwater flow

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ss

flo

w[k

g/s

]

0 60 120 180 240 300 3600

5

10

15

LP steam mass flow

Time [min]

Pre

ss

ure

[ba

r]

0 60 120 180 240 300 3602

3

4

5

LP drum pressure

Time [min]

Ma

ss

flo

w[k

g/s

]

0 60 120 180 240 300 3600

40

80

120

LP feedwater flow

Time [m in]

Te

mp

era

ture

[°C

]

0 100 200 300 400 500 600 700125

225

325

LP steam tem perature

Time [m in]

Pre

ss

ure

[ba

r]

0 100 200 300 400 500 600 7002

3

4

5

LP steam pressure

Time [m in]

Ma

ss

flo

w[k

g/s

]

0 100 200 300 400 500 600 7000

5

10

15

LP steam m ass flow

Time [m in]

Pre

ss

ure

[ba

r]

2

3

4

5

6

LP drum pressure

T ime [m in]

Ma

ss

flo

w[k

g/s

]

0 100 200 300 400 500 600 7000

50

100

150

LP feedwater flow

Tim e [m in]

Pre

ssu

re[b

ar]

0

10

20

30

IP drum pressure

Tim e [m in]

Ma

ss

flo

w[k

g/s

]

0 100 200 300 400 500 600 7000

5

10

15

IP feedw ater flow

Time [min]

Te

mp

era

ture

[°C

]

0 100 200 300 400 500 600 700125

225

325

IP steam tem perature

Time [min]

Pre

ss

ure

[ba

r]

0 100 200 300 400 500 600 7000

10

20

30

IP steam pressure

Tim e [min]

Ma

ss

flo

w[k

g/s

]

0 100 200 300 400 500 600 7000

4

8

12

IP steam m ass flow

Figure 16: IP steam parameters


Combined Cycle Start-Up Procedures: Dynamic Simulation and ...

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