Fast start-up analyses for Benson heat recovery steam generator

Energy 46 (2012) 295e309

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Fast start-up analyses for Benson heat recovery steam generator Falah Alobaid a, *, Stefan Pfeiffer a, Bernd Epple a, Chil-Yeong Seon b, Hyun-Gee Kim b a b

TU Darmstadt, Institut Energiesysteme und Energietechnik, Petersenstrasse 30, 64287 Darmstadt, Germany DOOSAN Heavy Industries and Construction, Thermal and Fluid Research Team, Changwon, South Korea

a r t i c l e i n f o

a b s t r a c t

Article history: Received 13 February 2012 Received in revised form 13 August 2012 Accepted 14 August 2012 Available online 19 September 2012

In this study, an entire simulation model is developed to investigate the influence of fast gas turbine start-up (within 20 min) on the dynamic behavior of a Benson heat recovery steam generator. The model includes a new high pressure feedwater control concept that considers a large number of live plant parameters and derivative functions. The advanced feedwater control circuit offers the possibility to operate the high pressure system in level mode (nature circulation) or Benson mode (once-through). Furthermore, it provides a rapid response and ensures a save operation to Benson HRSG (heat recovery steam generator) during the GT (gas turbine) abrupt transients and fast start-up procedures. A set of enhanced steam bypass control circuits are also implemented in the model. The developed HRSG model and its advanced control circuits are validated toward the design data at different part loads. The obtained results demonstrate very good physical agreement with relative error less than 3% and cite as evidence that the Benson model represents the real power plant. Finally, different fast start-up procedures (hot, warm, cold) are assessed with the validated Benson model, which provides adaptable operation demands to meet the gas turbine fast start-up requirements. Ó 2012 Elsevier Ltd. All rights reserved.

Keywords: Benson heat recovery steam generator Once-through Advanced control circuits Load changes Dynamic simulation Fast start-up procedures

1. Introduction The rapid worldwide increase in the consumption of fossil fuels in the last few decades suggests that the turning point for the depletion of the petroleum, natural gas and coal reserves has already been reached. However, the economic and social development of each country is highly depended on the energy security. As a consequence of this, the developing of new strategies for the thermal power plants to have higher performance and efficiency are assigned duties [1,2]. Modern combined cycle power plants (CCPP) among other thermal power plants are characterized by their high power generation efficiency, their low emissions and their high flexibility in mode of operation concerning rate of start-ups, load change times and shutdown gradients [3]. In a CCPP, the steam generator, that is arranged downstream of the gas turbine (GT), absorbs the waste heat of GT to generate steam and feed it into the steam turbine (ST). Typically, the CCPP can be fired by different fuels such as gas and oil. It can extend its fuel range to cover biomass and coal through the application of the new technology integrated gasification combined cycle (IGCC). On principle, a combined cycle power plant is operated at its design conditions. However, it also operates on the so-called off* Corresponding author. Tel.: þ49 6151 16 6691; fax: þ49 6151 16 5685. E-mail address: [email protected] (F. Alobaid). 0360-5442/$ e see front matter Ó 2012 Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.energy.2012.08.020

design load conditions due to the variation in consumer load. Recently, an increasingly important utilization for the combined cycle power plants is the compensation of varying electricity feedin from renewable energy sources like wind power. Dynamic changes of wind supply make the amount of power that provided to a network highly fluctuated. The CCPP can integrate with the wind power plants in order to overcome the negative dynamic impacts caused by the wind speed changes [4,5]. Independent of hot, warm- or cold start-up, new gas turbines reach 100% of their nominal load in approximately 20 min, while 70% of the nominal flue gas temperature and 60% of the nominal flue gas mass flow are already achieved approximately 7 min after the GT start [6]. The heat recovery steam generator (HRSG) should not hinder the operation as well as the abrupt transients of these gas turbines and thus should be designed in order to fully enable the GT fast start-up (within 20 min). Such gas turbines necessitate an accurate dynamic calculation of the HRSG that reaches its nominal load 50, 90 or 170 min after the GT start, depending on whether a hot, warm, or cold start-up procedure is performed (see Section 6). Benson HRSGs ensure the modern heavy duty gas turbine requirements and offer effective measures for achieving high efficiency. The Benson HRSGs do not have a high pressure drum and can follow the load change velocity of GT with better dynamic characteristics. They are designed for daily start-stop operation and can be run with steep start-up curves owing to the absence of the

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F. Alobaid et al. / Energy 46 (2012) 295e309

Nomenclature D E F f h h0 href k K M m _ m p Q R T t u z

a G g h v

r c

diameter [m] rate of entrainment [e] force/volume [N/m3] friction factor [e] static enthalpy [kJ/kg] stagnation enthalpy [kJ/kg] reference enthalpy [kJ/kg] loss coefficient over turbine section [-] Stodola coefficient [e] or metal factor [kW/ C] molar mass [kg/kmol] mass [kg] mass flow rate [kg/s] pressure [bar] heat flow/volume [kW/m3] rate of stratification [e] temperature [ C] time [s] longitudinal fluid velocity [m/s] longitudinal distance [m] interfacial heat transfer coefficient [kg/m3 s] or steam quality [e] mass transfer [kg/(m3 s)] isentropic exponent [e] polytropic expansion efficiency [e] specific volume [m3/kg] density [kg/m3] void fraction [e]

Subscripts a annular b bubbly d droplet fl form loss g gas phase grav gravitation i interface between phases k phase (either liquid or gas) l liquid phase s stratified flow or steam sat saturated sp specific ns non-stratified flow pu pump va valve w wall

high pressure drum. The HP drum greatly limits the allowable temperature gradients by the reason of its large wall thickness [7]. Faster plant start-ups are equivalent to fuel savings and can directly impact on CO2 emissions per MWth. With almost unlimited numbers of fast start-ups, the Benson HRSG has the advantage against the conventional drum HRSG with a limit of around 1000 fast start-ups due to the thermal stresses in thick wall of the high pressure components [3]. The dynamic simulation of the HRSG as a design and optimization tool for setting up the control concepts and selecting the operation processes gains nowadays more in importance. Validated HRSG models allow the designers to define the optimum plant performance, start-up procedures and the control circuits. In the literature, several steady-state works can be found regarding the

Abbreviations APROS advanced process simulation software Attemp attemperator avr average BFP boiler feedwater pump CPH condensate preheater CCPP combined cycle power plant DC device control ECO economizer EV evaporator FG flue gas funct function FW feedwater GT gas turbine Fin final HRSG heat recovery steam generator HPBPCV high pressure bypass control valve HPMSCV high pressure main steam control valve HP high pressure HT high temperature HRH hot reheater IPBPCV intermediate pressure bypass control valve IGCC integrated gasification combined cycle IPMSCV intermediate pressure main steam control valve Int interstage IP intermediate pressure IT intermediate temperature LPBPCV low pressure bypass control valve Ld plant load LP low pressure LPMSCV low pressure main steam control valve LT low temperature max maximal operator Mem memory min minimal operator PI proportional-integral controller Pr Prandtl number Re Reynolds number RH reheater RO density Select selector function SH superheater ST steam turbine Sub sub-cooling Sup superheating WS water/steam

thermodynamic analysis and process optimization of HRSG (among others: [8e10]). On the other hand, very few papers have dealt with the dynamic simulation of heat recovery steam generators, despite the significance of these investigations. On the prediction of the dynamic behavior of the single or dual pressure HRSG, the following works have been published [11e14]. The dynamic simulation of a triple pressure sub-critical HRSG during the start-up procedures has also been studied [15e17]. Recently, Alobaid et al. (2009) presented the upgrading of a validated sub-critical triple pressure HRSG dynamic model to a supercritical once-through HRSG up to 250 bar in the high pressure circuit [18]. However in the literature and according to the author’s knowledge, there are no studies up to date that have been dealing with the dynamic behavior of a sub-critical once-through HRSG (Benson) during the

F. Alobaid et al. / Energy 46 (2012) 295e309

297

Fig. 1. Flow diagram of the combined cycle power plant with Benson HRSG (simplified).

fast start-up. In this work, therefore, an entire dynamic sub-critical Benson HRSG model is built using the commercial simulation program “advanced process simulation software” (APROS). The model contains a modern high pressure feedwater control circuit that includes a large number of derivative functions such as the degree of sub-cooling at the evaporator inlet, the degree of superheating at the evaporator outlet, the derivative elements for considering the fluid side mass storage processes in the economizer and evaporator heat exchange surfaces etc. In addition, a set of enhanced feedwater and steam bypass control circuits are installed that enable the Benson HRSG model to simulate breakdown cases such as steam turbine trip, gas turbine load rejection and the fast start-up procedures. The generated Benson model and its improved control circuits are verified against the given data in steady state at different part loads. In conclusion, diverse breakdown cases and fast hot, warm and cold start-up processes (within 20 min) are performed with the validated Benson HRSG.

stream. The Benson HRSG has three pressure systems and a reheater section after the high pressure turbine, in detail as follows: 1. Once-through evaporator path in the high pressure circuit. 2. Natural circulation evaporator path in the intermediate and low pressure circuits. 3. Reheater section after the high pressure turbine. 4. Three steam turbines stages HP, IP, and LP. A schematic flow diagram of the combined cycle power plant including the gas turbine (G), the flue gas path (E), the three pressure circuits of the HRSG (A, B, C and D), the steam turbine (F), the feedwater and the steam bypass systems is shown in Fig. 1. The characteristic technical data of the real power plant for the gas and water/steam sides is presented in Table 1.

3. Mathematical model 2. Power plant description The combined cycle power plant comprises of a gas turbine connected to an unfired, horizontally based heat recovery steam generator. The gas turbine can be operated with gas or oil fuel. The exhaust gas passes the heat exchangers of the HRSG, where each heat exchanger section causes a pressure drop for the flue gas

The HRSG is modeled with the commercial software package APROS developed by VTT Finland [19]. APROS provides tools, solution algorithms and model libraries for full-scale modeling and simulation of the dynamic processes. The processes related component data are archived and maintained in different model libraries and based on non-linear equation systems that simulate

Table 1 Technical data of the real power plant. Benson HRSG

Steam conditions at turbine and condenser inlet HP

RH/IP

LP

Condenser

p ¼ 128.8 bar T ¼ 566.5  C _ ¼ 77:1 kg=s m

p ¼ 30.8 bar T ¼ 565.1  C _ ¼ 89:5 kg=s m

p ¼ 4.4 bar T ¼ 234.8  C _ ¼ 11:3 kg=s m

p ¼ 0.0056 bar T ¼ 35  C

Steam turbine thermal power

Flue gas conditions inlet/outlet

147 MW

_ ¼ 688:8 kg=s m TInlet ¼ 589.1  C TOutlet ¼ 87.7  C

298

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the real-life process status with a very high accuracy throughout the entire operating range of the plant. 3.1. Basic flow models

3.1.1. Homogenous flow model The one-dimensional homogeneous model is based on the mass, momentum and energy conservation equations of the mixture. Hence, water and steam have same velocities and temperatures in case of two-phase flow components. However, this flow model is suited for single-phase flow components including either water or steam such as superheater, reheater, turbine sections, economizers etc. Mass balance:

(1)

Momentum balance:


 vðruÞ v ru2 vp þ ¼ Fgrav þ Fw þ vt vz vz

(2)

Energy balance:

vðrh0 Þ vðruh0 Þ vp þ ¼ þ Qw vt vz vt

(4)

Momentum balance:

The Benson HRSG model includes different thermal hydraulic modules, i.e. the homogenous (three-equation or mixture) and the two-phase (six-equation or EulereEuler) flow models. Thermal hydraulics is described using the conservation equations for mass, momentum, energy and correlations for friction and heat transfer, as shortly described below according to [19e22].

vr vðruÞ ¼ 0 þ vz vt

Mass balance:

vðck rk Þ vðck rk uk Þ þ ¼ Gk vt vz

(3)

The terms Fgrav, Fw and Qw denote the gravitational acceleration force, the friction force and the heat flow through walls, respectively. In the energy equation, the term h0 is the total enthalpy (stagnation enthalpy) including the kinetic energy. The partial differential equations are discretized with respect to space and time and the non-linear terms are linearized. In the space discretization, the staggered discretization scheme has been applied. The state variables (such as pressure, enthalpy and density) are calculated in the middle of the mesh cells and the flow related variables (such as velocity) are calculated at the border of two mesh cells. For the enthalpy solution, the first order upwind scheme is utilized. In the mesh cell, the quantities are averaged over the whole mesh. The implicit method is applied for temporal discretization, where the pressure, void fraction and enthalpy linear equation groups are solved one after the other. The density is updated as a function of the solved pressure and enthalpy. 3.1.2. Two-phase model This flow model is suited for two-phase flow water/steam components, where there is a slip between the phases such as evaporator and condenser systems. The two-phase solution system of APROS is based on the one-dimensional conservation equations of mass, momentum, and energy for the two phases. The equations are coupled with empirical correlations describing various twophase phenomena, like friction and heat transfer for wall and interface. This means, the use of the two-phase (six-equation) model requires the knowledge of mass, momentum, and energy transfer between the phases. This transfer can be expressed from the flow parameters and their derivatives. When the equations are applied for the water and steam phases, a total of six partial differential equations are utilized.

  2 vðck rk uk Þ v ck rk uk vp ¼ Gk uik þ ck Fgrav;k þ Fwk þ Fik þ ck þ vz vz vt þ Fva þ Ffl þ DPpu ð5Þ Energy balance:



  v ck rk h0;k v ck rk uk h0;k vp þ ¼ ck þ Gk h0;ik þ Qik þ Qwk vt vt vz þ Fik uik

(6)

In these equations, the subscript k is either l ¼ liquid org ¼ gas. The subscripts i and w refer to interface between two phases and wall, respectively. The term G is the mass exchange rate between the phases. The symbols F and Q denote friction force and heat flow. The last three terms of the momentum equation are valve friction, friction from form loss and pump head. The interfacial heat transfer Qik is calculated separately for the liquid and gas phases as:


 Gas : Qig ¼ aig hg  hg;sat

(7)


 Liquid : Qil ¼ ail hl  hl;sat

(8)

where a represents the interfacial heat transfer coefficient and h is the static enthalpy. The correlations that are used to define the heat transfer coefficients between water and steam can be found by [21,22]. The determined interfacial heat transfers are applied as source terms to the energy equation (Equation (6)). For the wall heat transfer Qwk in Equation (6), there are separate heat transfer correlations for three heat transfer zones: wetted wall, dry wall and a transition zone between wetted wall and dry wall. If the wetted wall heat transfer is selected, only the liquid phase is assumed to be in contact with the wall of the flow channel. In the dry wall zone, only the gas phase touches the wall. The critical heat flux and minimum film boiling temperature are required for the selection of the heat transfer zone. The interfacial friction Fik in Equations (5) and (6) (the friction between the liquid and gas phases) is strongly dependent on the flow regime prevailing in the flow. Diverse interfacial friction correlations are applied to different flows. The modeled flow regimes are stratified flow and non-stratified flow consisting of bubbly, annular and droplet flow. The value for the interfacial friction is obtained as a weighted average of the different correlations. Void fraction c, rate of stratification R, and rate of entrainment E are used as weighting coefficients.

Fik ¼ RFis þ ð1  RÞFins

(9)

The interfacial friction Fins in non-stratified flow is:

Fins ¼ ð1  EÞ½ð1  cÞFib þ cFia  þ EFid

(10)

The interfacial friction Fis in stratified flow is:

Fis ¼

0:01½1 þ 75ð1  cÞrg DujDuj D

(11)

The forces Fia, Fib and Fid correspond to the interfacial friction in annular, bubbly and droplet flows, respectively [23]. The friction

F. Alobaid et al. / Energy 46 (2012) 295e309

Fwk between single-phase (liquid or gas) and the wall of the flow channel in Equation (5) is calculated with the formula:

2fk rk uk juk j D

Fwk ¼

(12)

The quantity fk is the friction pressure loss coefficient or friction factor for phase k. 3.1.3. Steam turbine simulation The steam turbine is simulated with the aid of momentum and energy equations (see Equations (2) and (3)), due to the fact that the homogenous flow model is used. The pressure and enthalpy drops are added as source terms in both equations. The pressure drop in steam turbine is calculated as:

1 2

Dp ¼ kru2

(13)

Whereas k is the loss coefficient in APROS:

2

  k ¼ f ðKÞ ¼  p K 2r 1 þ 2 p1

(14)

h

Where z ¼

g and g is the isentropic exponent. According to the g1

definition of the expansion efficiency in APROS, the process is polytropic. Therefore, the following equation can be derived:

h¼

1 vh v vp

(17)

If the specific volume v is solved from Equation (17) and introduced in Equation (16), then the Equation (18) is obtained.

h vp vh ¼ z p h

(18)

By integrating the right side term of Equation (18) from h2 to h1 and left side from p2 to p1, Equation (19).

h1 ¼ h2



p1 p2

h z



p1 p2

h z

0h2s  href ¼



h1s  href

p hz 2

p1 (20)

The enthalpy drop over a turbine section, which is based on assumption of the ideal gas, is determined using different models. When the fluid is steam, the enthalpy drop over a turbine is calculated as:

DHs ¼

  i h _ h1s  href  h2s  href m

(21)

h





i

_ DHs ¼ h1s  h1s  href ðp2 =p1 Þh=4:27 href m

(22)

is obtained. In Equation (22), the subscript s refers to steam, 1 to the state before the turbine and 2 to the state after the turbine. The quantity h is the polytropic expansion efficiency and href represents the reference enthalpy (1950 kJ/kg).

3.2. Benson HRSG model

(16)

z

h1s  href ¼ h2s  href

(15)

_ is the mass flow, p1 the pressure before the turbine,p2 the Where m pressure after the turbine and v1 represents the specific volume before the turbine. The enthalpy drop over the steam turbine is determined through the expansion efficiency. The basic expansion in the steam turbines is isentropic. Generally, the processes are not pure isentropic or isothermal but something in between. Therefore, the name of the process depends on the size of exponent. Hence, by giving different values for the efficiency, all processes from isothermal to isentropic will be covered. The expansion equation is obtained from Johanson’s equation. It is the equation of state for ideal steam, i.e. an ideal gas equation for steam using the static enthalpy instead of temperature [19,20].

pv ¼

is obtained. The variable z is set equal to 4.27 in the turbine section. Since the tabulated steam enthalpy is the water enthalpy plus the latent heat (evaporation) and the water enthalpy is set to 0 at T ¼ 0  C, the tabulated enthalpy cannot be applied to the ideal gas state. It has to be shifted with the reference enthalpy:

Equation (20) introduced in Equation (21), the Equation (22)

The symbols r and u denote to the density and velocity of fluid, respectively. The loss coefficient k is assumed to be constant and is calculated at the nominal load. The pressure loss over a turbine section or over a whole turbine block K is defined with Stodola Equation (15) [24].

qffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi
 _ p1 v1 = p21  p22 K ¼ m

299

(19)

The Benson heat recovery steam generator is built with different nets (19 in total) in order to have a high level of order and flexibility. The HRSG model has three pressure circuits up to 130 bar in the high pressure system and a reheater section after the HP turbine. The model includes the water/steam and the flue gas sides at a high level of detail, while the gas turbine section is considered as an inlet boundary condition with the characteristic parameters temperature, mass flow, pressure and flue gas composition. The flue gas path is modeled from the gas turbine exit to the HRSG exit. Hereby, each heat exchanger section causes a pressure drop of the flue gas stream. On the water/steam side, all bundle heat exchangers (45 in total) namely superheaters, reheaters, evaporators, and economizers have been implemented with real geometry data. Furthermore, for each pressure circuit: turbines, drums, pumps, condenser and valves are modeled, either. In this study, the heat exchangers are of eminent interest. Therefore, the high thermal stressed heat exchangers (superheater and reheater) are divided in sections as they are located in the high temperature region of flue gas path. The purpose of this is to get the exact temperatures of each section for their structural integrity. The overall heat absorbed was divided by the sections depending on its position in the flue gas path (the closer to the gas turbine outlet, the merrier heat can be absorbed from the flue gas by these sections). The heat transfer coefficient between flue gas and heat exchangers has an important influence on the heat flux and thus on the dynamic behavior of heat recovery steam generator. For the finned-tubes heat exchangers, there are a lot of heat transfer correlations available such as Schmidt, VDI and traditional ESCOA and revised ESCOA [6]. However, the application of one correlation is only possible for a definition region for Re and Pr-numbers as well as a tube geometry (e.g. fin pitch, bare tube diameter, number of fins per meter). In Benson HRSG model, the selection of the heat transfer correlation is depending on analyzed operation mode as well as the heat exchanger geometry [19e21].

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3.2.1. Model nets As examples for the nets of Benson HRSG model, the high pressure and the low pressure systems are shown in Figs. 2 and 3, respectively. The high pressure circuit is modeled from the outlet of the high pressure boiler feedwater pump (HP BFP) to the last superheater outlet (SH HP3) (see Fig. 2). The feedwater mass flow, which is streaming into the high pressure circuit, is delivered by the HP BFP. The feedwater flows into the HP evaporators via the high pressure economizers. The water in the HP once-through evaporators is heated by the flue gas and converted into steam. Subsequently, the steam exits the high pressure evaporators and streams into the superheater system, where the steam absorbs additional heat from the flue gas. The superheated steam exits the Benson HRSG and enters the high pressure turbine section. Interstage and final attemperators are provided at the inlet and outlet of the last superheater to control the temperature at the inlet of the HP turbine. Both attemperators will use the water from the high pressure feedwater pump. The feedwater mass flow is controlled by a feedwater control valve located in the same net. The high pressure economizer II (ECO HP5A, ECO HP5B, ECO HP6 and ECO HP7) is partly bypassed during the part load operation in order to prevent the high pressure economizer III (ECO HP8) from boiling. The bypassed mass flow is varied to keep the outlet temperature of the high pressure economizer III below the boiling point. During oil operation, the high pressure economizer I (ECO HP1, ECO HP2, ECO HP3, ECO HP4A, ECO HP4B and ECO HP4C) is partially bypassed so that economizer II absorbs more heat as the result of the lower inlet

temperature. Due to the decreased heat input into the IP evaporators, less amount of the IP steam is generated. Consequently, the intermediate pressure mass flow is reduced together with the increased pressure and the decreased steam in the low pressure drum. In view of the available overplus heat for the condensate preheater, the condensate inlet temperature therefore can be increased to the setpoint value avoiding the dew point corrosion on the low pressure economizers’ surfaces. The low pressure circuit that is divided into main and sub circuits is modeled from the inlet of the LP economizers to the LP superheater outlet (see Fig. 3). In the main circuit, the low pressure feedwater that is fed to the LP drum is delivered by the condensate preheater system. A level control valve is located between the last low pressure economizer outlet (ECO LP4B) and the low pressure drum. The water in the LP drum circulates through the LP evaporator’s tube bundle, where it is heated by the gas turbine exhaust gas and converted into saturated steam in the LP drum. The condensate preheated water at the inlet of the low pressure drum is fed to the feedwater pumps of the HP and IP circuits. The dry steam exits the LP drum and flows through the low pressure superheater, where the steam absorbs additional heat from the flue gas. After leaving the SH LP heat exchanger, the superheated steam enters the low pressure turbine. Beside the above mentioned LP path (main circuit), there is another path called the condensate preheater system (sub circuit). The condensate preheater system consists of several low pressure economizers namely (ECO LP1, ECO LP2, ECO LP3A, ECO LP3B, ECO LP4A and ECO LP4B). The main task of the condensate preheater is to feed the HRSG with the required mass

Fig. 2. High pressure net of the Benson HRSG model.

F. Alobaid et al. / Energy 46 (2012) 295e309

ECO LP3B

ECO LP2

ECO LP3A

M

301

From condenser

ECO LP1

Fuel preheating

CPH condensate recirculation

M

ECO LP4A

ECO LP4B

CPH cold bypass

LP drum LP process steam

M

To flash tank

CPH feedwater recirculation

GT air-preheating

SH LP EV LP To LP turbine

To HP economisers

To LP economisers HP feedwater pump

IP feedwater pump

Fig. 3. Low pressure net of the Benson HRSG model.

flow at a predefined temperature. Additionally, the HP BFP and the IP BFP are located in this net as both take out the feedwater from the condensate preheater and deliver it into the high and intermediate pressure economizers. 3.2.2. Model control circuits In this study, significant work has been achieved for building, improving and verifying the control circuits that must control all basic functions in the Benson HRSG. Applying the correct control concepts and choosing the realistic parameters of the controllers are fundamental for getting the optimal response of those control circuits during the dynamic behavior. To clarify the control concepts in the Benson HRSG model, the high pressure feedwater, the intermediate pressure drum level, the high pressure and the reheater bypass control circuits are explained in detail. The HP feedwater control loop depicted in Fig. 4 is the most complex arrangement in the Benson HRSG. It provides the main feedwater setpoint and offers the possibility to operate the high pressure circuit in different modes, namely the separator level mode (natural circulation) and Benson operation mode (oncethrough). The basic idea behind the feedwater mass flow controller is derived from the available heat that can be absorbed by the HP evaporators from the flue gas (DhFG EV) and the medium enthalpy setpoint at the inlet of the HP separator (DhWS EV). This signal is corrected by the derivative element for considering additional heat output of the metal masses of the evaporator (Kmetal EV), by the derivative element for considering fluid side mass storage processes in the economizer heat exchange surfaces (DrWS ECO), by the degree of sub-cooling at the evaporator inlet (Dsub), by the

degree of superheating (Benson mode) or the steam quality (level mode) at the evaporator outlet (Dsup,a) and by the attemperators mass flows as function of the feedwater mass flow. Further correction is considered depending on the operation mode (level or Benson). While in the separator level mode, the water level within the HP separator should be kept at a fixed level independent from the HRSG load, in the Benson mode the enthalpy at the outlet of the HP separator should meet the load dependent enthalpy setpoint. The selection between the level and Benson modes is related to the load condition of the HRSG. The level mode is switched over to the Benson mode, if the HRSG load exceeds 30%. The control circuit of the intermediate pressure drum (see Fig. 5) is three element drum level control. It regulates the IP feedwater mass flow by controlling the IP feedwater valve, which is located downstream of the intermediate pressure economizers. The operation algorithm of the IP drum level controller is described as follows: on the one hand, the difference between _ is the IP feedwater mass flow and the IP steam mass flow dm measured. On the other hand, the difference between the IP drum level setpoint and the actual value dL is measured. The _ dL) affects a PI-controller. The deviation of these two paths (dm, PI-controller commands the continuous device control (DC) that operates the IP drum level valve. If the pressure in the IP drum increases over a certain value (maximum pressure), the deviation _ dL) will be replaced by the pressure difference between the (dm, IP drum and the maximum pressure setpoint in order to prevent the further raise in the IP drum pressure. Fig. 6 shows the high pressure bypass control circuit that routes the high pressure steam, which is not accepted by the high pressure

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F. Alobaid et al. / Energy 46 (2012) 295e309

+

lECO f(t,lECO) vavr

/

PI-Entalphy

. . f(mFW,mFWsetpoint) Ld

x m. entalphy setpoint +

PI-level +

f(Tsat,T) ,

+ . mlevel + setpoint

PI

sub f(Tsat,T) ,

Tsetpoint

DC

EV 1-10 0

ECO O 1-8 18

A

f(Level,Entalphy)

. m' ECO

L

Lsetpoint

. msetpoint

+

h' 1WS f(p,T-T f(p,TTT

. m' sub,sup +

+

Tfin Tint

HP BFP

. mFW -

sup

h2WS f(p,T)

Seperator

+

f(T) roavr

x

Fin Attemp

Int Attemp

funct

Selector

. m' level

A A

h' 2WS

/

. mFW

[0-1]

+ -

x

VECO

min

PI-Attemp + h'attemp h' sup

-

h' 2WS f(p,T+T sup) Kmetal

/

. msetpoint

T HP turbine To

dTSat

+

sub)

+

SH S H 1-3

x

EV

. mFG

x + h'2FG f(p,T+Tpinch)

h1FG f(p,T)

+

Gas turbine outlet

Stack FG ECO

FG SH

FG EV VAP

Fig. 4. HP FW control circuit (simplified).

turbine into the cold reheater. Generally, it works during the Benson HRSG start-up, as long as the HP steam quality has not matched the high pressure turbine requirements. Furthermore, the high pressure bypass system is used during the steam turbine trip at any load. An Attemperator has been implemented, which cools the steam behind the HP bypass control valve to 50  C above the

saturated steam temperature before it enters the cold reheater system. The injection water is delivered from the high pressure feedwater pump. The main tasks of the HP bypass controller are to insure a smooth build-up of the high pressure during the start-up (see Fig. 7) and to prevent the high pressure from decreasing during the turbine trip. The principle of the work:

pmax setpoint + To IP turbine

p IP BFP min

PI mFW

ECO 1-2

IP drum

SH

L

DC

EV 1-3 + Delta L +

Delta m

+ mSH steam

Fig. 5. IP FW control circuit.

+ Lsetpoint

F. Alobaid et al. / Energy 46 (2012) 295e309

SH HP3 psetpoint - Minimum pressure setpoint + PI

mSH p

PI

-

mSH steam DC Time-position control function

Automatic setpoint adjuster

+ +

HPMSCV HP turbine

mSH

min

+

303

HPBPCV Reheater

- 1 DC

psetpoint

Fixed pressure setpoint min

funct

Throttling function

Mem

+

-

+

Selector

Steam turbine trip

PI

Attemp

PI -

HP BFP

GT load rejection

PI

psetpoint calculation +

Standing idle pressure

PI

pshut-down setpoint + mSH steam setpoint -

+

PI -

Starting steam turbine

mRH Fig. 6. HP bypass control circuit.

I. During the start-up, both HP bypass valve (HPBPCV) and HP main steam valve (HPMSCV) are initially closed. This procedure enables the pressure in the high pressure circuit to increase rapidly. When the minimum pressure setpoint is met, the HP bypass valve starts opening to counteract further rising of the pressure. Due to the continuous pressure increasing, the HPBPCV is forward opened (see Fig. 7).

II. After the HP bypass valve is fully open, the steam pressure rises due to the steam generation increasing. At this stage, the automatic setpoint adjuster is activated. It tracks the bypassed pressure at interval about 1 bar below the high pressure. III. When a specified high pressure is reached, the HPBPCV is throttled to intermediate start position (circa 70%). This

Fig. 7. HP pressure, HPMSCV and HPBPCV characteristics during start-up.

304

F. Alobaid et al. / Energy 46 (2012) 295e309

IPMSCV

p

mHRH

RH IP2 (HRH) p mHRH -

IP turbine mRH steam DC

pressure setpoint

+

Time-position control function

PI

psetpoint

IP drum max pressure setpoint Feedforward control PI

-

IPBPCV Condenser

+

PI Starting steam turbine

-

-

mRH steam setpoint

Attemp

+

min

IP drum pressure

+

DC

mCondenser

Mem

-

Steam turbine trip PI

+ - p setpoint calculation

Selector

+

PI

IP BFP pshut-down setpoint

PI Standing idle pressure +

GT load rejection

-

Fig. 8. RH bypass control circuit.

process ensures the high pressure to reach the fixed pressure setpoint faster. After the fixed pressure is met, the HP bypass valve limitation is switched off to hold the pressure at the fixed value by further opening of the HPBPCV (see Fig. 7). IV. When the main steam control valve starts opening, the HP bypass valve starts closing in the same attitude. If the HPMSCV is fully opened, then the HPBPCV should be closed. The high pressure bypass valve could be again opened, if the pressure in the high pressure circuit is increased over a defined setpoint (the nominal pressure pulse delta p) in order to discharge the high pressure. V. In case of the steam turbine trip, the HP bypass control circuit is put in operation. While the HPMSCV is immediately closed, the HPBPCV controls the high pressure to reach the pressure level, which was existent before the steam turbine trip. Holding the pressure at high level has the advantage that the HRSG is already prepared to the hot restart. VI. In case of gas turbine load rejection, the gas turbine exhaust temperature drops very fast to low level. To prevent the condensation in the high pressure superheater, the HPBPCV will open to reduce the high pressure to its saturation temperature below the gas turbine exhaust temperature.

Table 2 SH HP outlet at different part loads.

In Fig. 8, the RH bypass control circuit is presented. The intermediate pressure main steam valve (IPMSCV) adjusts the mass flow to the IP turbine. It is controlled by a time gradient function in the same way as the HPMSCV. During the normal operation, this valve is opened while at start-up, it is closed. The intermediate pressure bypass valve (IPBPCV) is controlled by a summation rule. Herein two different control goals are detected. On the one hand, the remainder mass flow of the IPMSCV has to be bypassed. On the other hand, the reheated pressure has to be controlled at the beginning on a designated setpoint value but should not exceed a maximum value. Furthermore an attemperator has been implemented in order to cool down the steam behind the control valve to 50  C above saturated steam temperature. The attemperator uses water from the condensate pump. When breakdown cases occur such as steam turbine trip and gas turbine load rejection, then the reheater bypass system is operated like the high pressure bypass control circuit. 4. Model validation In this section, the Benson HRSG model will be validated at 100%, 78%, 50% and 25% part loads. The exhaust mass flow and

Table 3 RH outlet at different part loads. RH steam mass flow [kg/s]

HP SH outlet

HP SH steam mass flow [kg/s]

HP superheated temperature [ C]

HP SH pressure [bar]

RH outlet

Load %

Design data

Simulated results

Design data

Simulated results

Design data

Simulated results

Load % Design data

100 78 50 25

77.1 70.6 e 35.5

77.1 70.1 58.8 35.8

566.5 566.3 e 502.7

565.1 567.1 567.2 502.5

128.8 118.6 e 62.1

129.7 119.6 110.9 61.4

100 78 50 25

89.5 84.5 e 41.1

RH superheated temperature [ C]

RH pressure [bar]

Simulated results

Design data

Simulated results

Design data

Simulated results

89.6 83.6 70.1 42.1

565.1 565.1 e 501.7

564 564.6 564.6 500.4

30.8 29.1 e 14.3

30.4 28.5 23.9 14.1

F. Alobaid et al. / Energy 46 (2012) 295e309

305

Table 4 SH LP outlet at different part load. LP SH outlet

LP SH steam mass flow [kg/s]

LP superheated temperature [ C]

LP SH pressure [bar]

CPH FW mass flow [kg/s]

Load %

Design data

Simulated results

Design data

Simulated results

Design data

Simulated results

Design data

Simulated results

100 78 50 25

11.3 9.5 e 2.9

11.3 9.4 6.7 2.9

234.8 232 e 198.4

232.9 228.9 220.7 197.5

4.4 4.1 e 1.8

4.3 3.9 3.2 1.8

103.9 96.7 e 61

103.9 96.9 82.5 62.5

FG_Mass flow

5. Steam turbine trip Among a variety of possible breakdown cases that can be taken place at the combined cycle power plant, the steam turbine trip is engaged in particular importance. By reason of steam turbine protection, the isolation of steam turbine must be achieved within seconds. Hence, the steam should be not routed anymore into the HP, IP and LP turbine stages. The best solution is carried out by bypassing the steam into the condenser through the HP, RH and the LP bypass systems. The bypass control circuits adjust the pressure to meet the pressures, which were existent before the steam turbine trip to enable the hot restart. In Fig. 9, the boundary condition for the steam turbine trip is displayed (see Appendix B). The HP, RH and LP steams are immediately bypassed the turbine stages, as it is previously mentioned. Accordingly, the power output of steam turbine falls down to zero. During the steam turbine trip, the behavior of the Benson HRSG circuits is almost the same, so the dynamic behavior of HP system will be only presented (see Fig. 10). The high pressure main steam valve closes immediately to protect the HP turbine. With the aid of a memory component (see Fig. 6), the high pressure level before the turbine trip is saved as setpoint. The high pressure bypass valve will close very fast in order to hold the pressure at its setpoint. For this reason, the HP steam mass flow drops rapidly to zero. The superheated temperature behaves almost qualitatively to the flue gas boundary condition and reaches 300  C after 250 min. The steam mass flow and temperature peaks in the period of time between the 40 and 100 min result from the increasing of high pressure over the setpoint. In time range between 100 and 150 min, the high pressure falls below the pressure setpoint so that the high pressure bypass valve closes completely allowing the pressure to be increased. After 200 min

Steam Power

Temperature

Steam mass flow

Pressure

150

580

150

600

120

518

120

450

90

456

90

300

60

394

60

150

30

332

30

0 0

50

100

150 Time [min]

200

250

0 300

Fig. 9. Dynamic boundary conditions during the steam turbine trip.

Temperature [°C]

750

Power [MW]

Mass flow [kg/s] Temperature [°C]

FG_Temp

9 kg/s. At 25% load, the simulated LP steam mass flow, the pressure and the temperature agree well the design data with about 2% relative error. Moreover, the CPH feedwater, which represents the main mass flow in the power plant, shows a high accuracy toward the design values with a maximum relative error less than 2.5% over the whole load range.

270 0

50

100

150 Time [min]

200

250

0 300

Fig. 10. HP SH dynamic behavior during the steam turbine trip.

Mass flow [kg/s] Pressure [bar]

temperature of gas turbine are applied as dynamic boundary conditions to the HRSG inlet (see Appendix A). The flue gas composition has been assumed to be constant over the entire load range of gas turbine. Table 2 displays the steam temperature, the pressure, and the steam mass flow at the outlet of HP superheater at different loads. The comparison between the model predictions and the design data shows high accuracy to each other. As mentioned previously, the Benson mode is switched over to level mode, if the HRSG load is less than 30%. In Benson mode, the HP superheated temperature is kept constant equal to 565  C independent of the load changes because the high pressure feedwater controller and the high pressure attemperators. Due to the decreasing of the flue gas temperature at 25% load, the HP temperature falls down to 502  C in both measurement and simulation (the level mode is activated). The simulated HP steam mass flow and the pressure match well the design data at different loads with maximum relative error of 1% and 2%, respectively. In Table 3, the comparison between the simulation results and the design data at the outlet of reheater section is presented. The comparison shows a good agreement between the simulations and the given data at different loads with relative error less than 2.5%. The simulated reheated steam mass flow and the temperature agree well with the design data. Similar to the high pressure circuit, the reheated temperature is also kept constant equal to 565  C at 78% and 50% loads because of the RH attemperators (the Benson mode is on). The reheated temperature and the pressure follow the dynamic transients of the flue gas mass flow and temperature at 25% load, where the reheated temperature decreases from 565  C to 500  C and the reheated pressure decreases to 14 bar. The values of the steam mass flow, the temperature and the pressure at the outlet of low pressure superheater at different loads are displayed in Table 4. In addition, the condensate preheater mass flow (CPH) is depicted. At 78% load, the LP temperature and pressure decrease from 233  C and 4.2 bar to 229  C and 4 bar, respectively. By the reason of flue gas mass flow and temperature reduction i.e. less available heat to be absorbed by the low pressure evaporators, the generated low pressure steam is decreased to

F. Alobaid et al. / Energy 46 (2012) 295e309 FG_Mass flow

Temperature

Steam Power

Pressure

600

128

506

100

450

96

432

75

300

64

358

50

150

32

284

25

0

20

40

60 Time [min]

80

210

0 120

100

0

20

from the turbine trip, the high pressure is almost constant (128 bar). Analogously to the high pressure, the intermediate and low pressure are kept constant at 30 bar and 4 bar by the reason of the RH and LP bypass control valves, respectively. Holding the pressure at its nominal values has the advantage that the Benson HRSG is already prepared to the hot restart. 6. Fast start-up analyses The HRSG model analyses with the design data for steady state, different load changes and breakdown cases have been performed with high accuracy. Finally, the fast start-up simulations with the validated model can be carried out. The exhaust flue gas mass flow and the exit temperature of gas turbine are applied as dynamic boundary conditions to the Benson model. The exhaust flue gas composition has been assumed as constant over the entire load range of gas turbine. Independently from the hot, warm- or cold start-up, the gas turbine reaches 100% of its nominal load in approximately 20 min. In the following sections, the influence of the transient behavior on the physical values (temperature, pressure and the mass flow) of the high, intermediate and low pressure circuits will be explained during the fast start-up in detail. 6.1. Hot start-up Figs. 11e13 display the flue gas boundary conditions, the dynamic behavior of superheater and reheater systems during the fast hot start-up. The superheated temperature increases from the standing idle temperature 265  C to the nominal temperature 565  C within 10 min (the thermal gradient is 30  C/min). From time point 36 min, the superheated temperature is kept constant at

Steam mass flow

60 Time [min]

80

0 120

100

Fig. 13. RH dynamic behavior during the fast hot start-up.

Fig. 11. Dynamic boundary conditions during the fast hot start-up.

Temperature

40

Mass flow [kg/s] Pressure [bar]

125

Temperature [°C]

580

0

its nominal value due to the influence of the high pressure feedwater control circuit. The HP FW controller adjusts the feedwater mass flow to hold the enthalpy at the outlet of HP evaporators at a predefined setpoint. The sudden increase of superheated temperature over 565  C will be prevented with the aid of the interstage and final attemperators. The high pressure takes 20 min to reach the fixed pressure value (80 bar) from its standing idle pressures (40 bar). In the period of time between 20 and 36 min, the high pressure is held at its fixed pressure setpoint (80 bar). From 37 min, the HRSG load increases further and the high pressure rises gradually to reach its maximum value of 129 bar. After 8 min from the gas turbine start-up, the high pressure evaporators start generating steam. The high pressure steam mass flow needs about 40 min as well to reach the nominal value (77 kg/s). The HP steam mass flow oscillates in the period of time between 20 and 50 min. These vibrations result from the changing over the level mode to Benson mode and the switching over the HP bypass valve to the HP main steam valve. Similar to superheated temperature, the reheated temperature rises rapidly during the fast hot start-up from 270  C to 565  C within 15 min (circa 19  C/min). The reheated temperature remains then constant due to the RH interstage and final attemperators. Within 50 min, the pressure increases gradually from 10 bar to reach its operation value. The reheated steam mass flow behaves almost like the superheated steam mass flow. During the fast hot start-up, the reheated steam mass flow reaches a maximum value of 100 kg/s about 10% higher than its operation value (90 kg/s). The reason of this overplus is coming from the HP bypass attemperator that cools down the high pressure steam before it enters the reheater section (see Section 3.2 and Fig. 6). In the period between 35 and 45 min, the high pressure bypass valve starts closing and the additional mass flow of the HP bypass attemperator will be

Pressure

FG_Temp

FG_Mass flow

Steam Power

150

750

150

506

120

600

120

432

90

450

90

358

60

300

60

284

30

150

30

210

0

20

40

60 Time [min]

80

100

0 120

Fig. 12. HP SH dynamic behavior during the fast hot start-up.

Mass flow [kg/s] Temperature [°C]

580

Mass flow [kg/s] Pressure [bar]

Temperature [°C]

Steam mass flow

160

Power [MW]

Mass flow [kg/s] Temperature [°C]

FG_Temp

750

0 0

20

40

60 80 Time [min]

100

120

0 140

Fig. 14. Dynamic boundary conditions during the fast warm start-up.

Steam Power [MW]

306

F. Alobaid et al. / Energy 46 (2012) 295e309 Pressure

FG_Temp

FG_Mass flow

Steam Power

750

150

498

120

600

120

416

90

450

90

334

60

300

60

252

30

150

30

170 0

20

40

60 80 Time [min]

100

120

Mass flow [kg/s] Temperature [°C]

150

0 140

0 0

30

60

90

120 150 Time [min]

180

210

Power [MW]

Steam mass flow

580

Mass flow [kg/s] Pressure [bar]

Temperature [°C]

Temperature

307

0 270

240

Fig. 15. HP SH dynamic behavior during the fast warm start-up.

Fig. 17. Dynamic boundary conditions during the fast cold start-up.

eliminated. The reheated steam mass flow starts first entering into the IP turbine after 20 min from the fast hot start-up, while the superheated steam to the HP turbine begins 5 min later. It takes about 20 and 25 min until the total reheated and the superheated steam mass flows are completely directed into the IP and HP turbines. As a sequence of this, the steam turbine reaches its nominal load 50 min after the GT start (see Fig. 12).

(circa 40  C/min) is about 25% faster than the hot start-up. From the point in time 30 min, the superheated temperature is kept constant at the same altitude. The high pressure increases from its standing idle pressures 16 bar to the minimum pressure setpoint (20 bar) within 5 min and to the fixed pressure setpoint (80 bar) within 24 min. In the distance between 25 and 50 min, the high pressure is hold at 80 bar. From 50 min, the high pressure rises gradually to reach its maximum value 128 bar. The simulated pressure characteristic during the fast warm start-up agrees very well with the required pressure curve in the Fig. 7. The reheater section shows similar behavior to the high pressure circuit (see Fig. 16). The reheated temperature rises rapidly to its operation value within 12 min (circa 27  C/min). The reheated pressure and the reheated steam mass flow behave qualitatively to the flue gas boundary condition. As it is previously explained, the peak in the reheated steam mass flow is caused by changing over from level to Benson mode. In the period of time between 40 and 45 min, the reheated steam mass flow reaches 103 kg/s, which is about 13% higher than its nominal value. This overplus in the reheated steam mass flow results from the HP bypass attemperator that cools down the bypassed steam flow to a predefined value before it is sent into the reheater section. The IPMSCV starts opening after 30 min from the fast warm start-up, while the HPMSCV starts 12 min later. From 42 min, both valves continue opening till they reach their 100% position within 48 min.

Fig. 14 shows the exhaust flue gas mass flow and the exit temperature of gas turbine during the fast warm start-up. The gas turbine is at full load within 20 min, while the Benson HRSG requires circa 90 min to reach its nominal operation. The behavior of the relevant parameters at the last high pressure superheater during the fast warm start-up procedure is displayed in Fig. 15. After 14 min from the GT start, the high pressure steam mass flow begins increasing (about 6 min later if it is compared to the fast hot start-up). It needs then circa 70 min to achieve the nominal value. The reduction in the HP steam mass flow in the period between 20 and 30 min is caused by switching over from level to Benson mode. In the Benson mode, a fixed superheating degree at the evaporator outlet (45  C above the saturated temperature) should be met, while in the level mode, the saturated temperature is dominated. It is clear that the generated steam mass flow in Benson mode to agree with the superheating criterion is smaller than the steam that is generated in the level mode. Other oscillations in the superheated steam mass flow can be detected between 30 and 60 min. The opening of HPMSCV and the closing of HPBPCV stand for these minor vibrations. Within 9 min, the high pressure temperature increases from its standing idle temperature (200  C) to the nominal temperature 565  C. This thermal gradient

Steam mass flow

In the fast cold start-up, the Benson model needs 170 min after the gas turbine start-up to reach its full load operation (see Fig. 17). In Fig. 18, the behavior of the relevant parameters at the last high pressure superheater is shown. The high pressure increases

Temperature

Pressure

Steam mass flow

Pressure

125

580

150

504

100

484

120

428

75

388

90

352

50

292

60

276

25

196

30

200 0

20

40

60 80 Time [min]

100

120

0 140

Fig. 16. RH dynamic behavior during the fast warm start-up.

Temperature [°C]

580

Mass flow [kg/s] Pressure [bar]

Temperature [°C]

Temperature

6.3. Cold start-up

100 0

30

60

90

120

150

180

210

240

0 270

Time [min] Fig. 18. HP SH dynamic behavior during the fast cold start-up.

Mass flow [kg/s] Pressure [bar]

6.2. Warm start-up

308

F. Alobaid et al. / Energy 46 (2012) 295e309

to HP turbine

to Cold Reheat

from RH

105

68

84 Mass flow [kg/s]

Mass flow [kg/s]

from SH HP

85

51 34 17

to IP turbine

to condenser

63 42 21

0 0

30

60

90

120 150 Time [min]

180

210

240

270

0 0

30

60

90

120

150

180

210

240

270

Time [min] Fig. 19. HP bypass dynamic behavior during the fast cold start-up. Fig. 21. RH bypass dynamic behavior during the fast cold start-up.

gradually from 1 bar to the minimum pressure setpoint of 20 bar within 10 min. For circa 5 min, the high pressure remains at the same value. The high pressure jumps then rapidly from 20 bar to the fixed pressure value (80 bar) within 10 min. In the period of time between 25 and 135 min, the high pressure is kept constant equal to 80 bar. Beginning from 135 min until 180 min, the high pressure starts increasing to reach the design value of 128 bar. The smooth build-up of the high pressure during the fast cold start-up matches well with the pressure trend that is aimed in Fig. 7. The HP steam temperature rises extremely from 120  C to 565 within 7 min (63  C/min). The temperature gradient is circa 36.5% faster than the warm start-up and 52% faster than the hot start-up. After 18 min from the start-up, the high pressure steam mass flow starts increasing (circa 10 min later compared to the fast hot start-up). It needs then about 130 min to achieve its nominal value. The oscillations in the HP steam mass flow result from changing over between level and Benson mode and switching over from HPBPCV to HPMSCV. In this section, the behavior of the high pressure bypass system during the fast cold start-up will be explained in detail (see Fig. 19). The HP main steam and the HP bypass valves are initially closed before the start-up in order to enable the fast increasing in the high pressure (see Figs. 7 and 18). From point of time 15 min, the high pressure reaches the minimal pressure setpoint and the HPBPCV starts opening. In sequence for this, the high pressure steam mass flow increases. In the period of time between 15 and 95 min, the high pressure steam streams into the reheater section after it is cooled down by the HP bypass attemperator. At 100 min, the HPMSCV starts opening and the HPBPCV starts closing in the same ratio. To warm-up the HP turbine, a break point can be detected between 105 and 130 min. After 170 min from the GT start-up, the

Steam mass flow

Pressure

125

484

100

388

75

292

50

196

25

100 0

30

60

90

120

150

180

210

240

0 270

Time [min] Fig. 20. RH dynamic behavior during the fast cold start-up.

Mass flow [kg/s] Pressure [bar]

Temperature [°C]

Temperature

580

HPBPCV is completely closed and the HMSCV is completely opened. Here, it should be mentioned that the HPMSCV is commanded by a time function depending on the steam turbine start-up requirements. The behavior of HP bypass system during the fast hot and warm start-ups can be predicted by the same token. Fig. 20 displays the behavior of the relevant parameters at the last reheater during the fast cold start-up procedure. The reheated temperature increases from the standing idle temperature 120  C to the maximum temperature 565 within 12 min. This equates to 40.5  C/min thermal gradient, which is twice as much during the fast hot start-up. The reheated pressure raises gradually to the design data. The reheated steam mass flow shows similar characteristic to the HP steam mass flow. The overplus in the steam mass flow appears in the period of time between 60 and 130 min. After 130 min from the GT start-up, the HP bypass valve starts closing and the HP bypass attemperator mass flow is set equal to zero. The behavior of RH bypass system is displayed in Fig. 21. During the cold start-up, the Benson model needs about 18 min to start generating steam. At this point, the reheated steam mass flow starts streaming into the condenser. From 45 min, the IPMSCV starts opening and the IPBPCV starts closing. For warming-up the IP steam turbine, two hold points are required. While the first one lasts 45 min, the second one takes 30 min. At 170 min, all the reheated steam is routed into the IP turbine. 7. Conclusion In this work, a numerical model is developed to understand the influence of fast gas turbine start-up on the dynamic behavior of a triple pressure sub-critical Benson heat recovery steam generator. The Benson HRSG model includes a new feedwater control circuit and a set of enhanced steam feedwater bypass systems. The developed model and its advanced control circuits were evaluated toward the design data at different part loads. The received results demonstrate a good agreement with a relative error less than 3%. The analyses of the Benson HRSG during the fast start-up processes (hot, warm and cold within 20 min) provide an evidence that Benson is able to meet the gas turbine fast start-up requirements. The results show that the Benson HRSG reaches its nominal load 50, 90 or 170 min after the GT start, depending on whether a hot, warm, or cold start-up procedure is performed. An important result was also obtained by the application and the verification of a new high pressure feedwater control circuit that considers a large number of plant parameters and derivative functions. The developed high pressure control circuit offers the possibility to operate the Benson model in different modes, namely level or once-through. In conclusion, the validated model can serve for further simulation

F. Alobaid et al. / Energy 46 (2012) 295e309

studies and prospective dynamic works in order to meet and improve the requirements of the real high efficiency combined cycle power plant. Hence, the strategy of model building up and the developed control circuits can be transferred to other applications by the same procedure. Appendix A A summary of the part load dynamic boundary conditions:         

100% full load steady state for 50 min. Load reduction from 100% load down to 78% load within 6 min. 78% load steady state for 166 min. Load reduction from 78% load to 25% load within 6 min. 25% load steady state for 166 min. Load increase from 25% load to 50% load within 6 min. 50% load steady state for 166 min. Load increase from 50% load to 100% load within 6 min. 100% full load steady state for 166 min.

Appendix B A summary of the dynamic boundary conditions during the steam turbine trip:  100% full load steady state for 45 min.  Shut down within 2 min.  Steady state simulation for 250 min.

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