Cogeneration Power-Desalting Plants Using Gas Turbine Combined Cycle

The gas-steam turbine combined cycle (GTCC) is the preferred power plant type because of its high efficiency and its use of cheap and clean natural gas as fuel. It is also the preferred type in the Arab Gulf countries where it is used as cogeneration power-desalting plant (CPDP). In this chapter, descriptions and analysis of the GTCC components are presented, namely, the gas turbine cycle (compressor, combustor, gas turbine), heat recovery steam generator, and steam turbine. Combinations of the GTCC with thermally driven desalination units to present CPDP are presented. A parametric study to show the effect of using GTCC on several operating parameters on the CPDP is also presented, as well as cost allocation methods of fuel between the two product utilities (electric power and desalted seawater are also presented).


Introduction
The efficiency of power plant (PP) using gas turbine (GT) combined cycle (GTCC) is higher than that of steam cycle PP prevailed as baseload-type plant before 2000, and the GT power cycle that was used as peak load and starting units in these steam PPs. The efficiency of the steam turbine (ST) plants is in the range of 35-40 %, the GT is in the range of 30-36 %, and GTCC is in the range of 45-58 % (Figure 1). This is the main reason for the GTCC to become the preferred-type PP in the Gulf Cooperation Countries (GCC) and worldwide, besides using clean and cheap natural gas (NG) as a fuel. Increasing the PP efficiency reduces the emissions of greenhouse gases (GHG), sulfur dioxide, and nitrogen oxide. The GTCC (Figure 2a) includes an upper GT cycle (i.e., compressor, combustion chamber (cc), gas turbine) and bottom steam cycle (i.e., heat recovery steam generator (HRSG), steam turbine (ST), and condenser). The GT cycle produces almost 2/3 of the GTCC electric power (EP) output, with mainly NG used as fuel supplied to the cc. Hot gases exhausted from the GT (typically at 475-600 ο C) are directed to HRSG to produce steam. This steam, when at high-enough temperature and pressure, is directed to ST to produce more work without adding more fuel. The ST cycle produces about 1/3 of the GTCC power output. Processed heat (in the form of steam) produced from the HRSG or extracted (or discharged) from the ST can be used to operate desalting plant (DP), district heating absorption cooling, and/or other processes.  [2] Most PPs in the GCC are cogeneration power-desalting plants producing both desalted seawater (DW) and EP in single plants called CPDP. The desalting units are supplied with its needed low-pressure (LP) steam by extracting (or discharging) steam from the ST of the GTCC or directly from the HRSG when the ST is not operated or does not exist. The ST used in the GTCC can be extraction-condensing steam turbine (ECST) or back-pressure steam turbine (BPST) discharging all of its steam to the DP. Examples of recently installed CPDP using GTCC are Shuaiba North in Kuwait, (Figure 2b), Jebel Ali in the UAE, and Ras Girtas and Mesaieed in Qatar.
In Kuwait, all power plants were of steam type before 2003 and were combined with mainly multistage flash (MSF) desalting plants (DP) up to 2003 to form CPDP; see Table 1. The use of ST for power production in Kuwait followed the 1980s general world trend of using ST in the PP, when the share of GTCC plants was very limited. In the 1990s, the share of GTCC increased very rapidly in the world due to extensive improvements in the GTs. These improvements in GT resulted in reliable GTCC technology and low capital cost of the GTCC plants compared to the ST cycle of the same capacity. NG availability at low cost in many parts of the world and the high efficiency of the GTCC (and thus the use of less fuel with less impact on the environment) promoted the share of the GTCC all over the world. Today, the GTCC-type PP becomes the preferred choice of PPs in most areas in the world, particularly in the GCC as shown in Table 2. Moreover, the GTCC equipment costs are less than that of the conventional ST plants.  This article presents description of GTCC power plant units, detailing the GT, HRSG, ST, and their combining arrangements with DP to form CPDP. Methods of allocating the fuel supplied to the CPDP using GTCC between the EP and DW are presented.

The gas turbine cycle
In the past, STs were the prevailing type of PPs used to satisfy the baseload. GTs were included in these PPs to satisfy peak load and daily operation at time of high power demand lasting few hours in summer months and as emergency system. GTs for peak-load applications operate for short periods, few hours, e.g., 2-500 h/y, with no concern to thermal efficiency, but fast loading and start reliability are the concerns. Emergency GT units have to reach full load in a very short time; and aeroderivative GTs were originally designed to be capable of producing full power from cold metal in 120 seconds. The GT is usually classified as heavy frame industrial and aeroderivative types. Aeroderivative gas turbines use advanced aircraft engine to provide flexible, lightweight, and compact GTs. Heavy frame-type GTs are usually slower in speed, narrower in operating speed range, heavier, larger, have higher air flow, and slower in start-up. Traditionally, preference has been to place heavy frame industrial units in easily accessible baseload applications.
Developments of GT in terms of high unit capacity (up to 400 MW; see Figure 3a), reliability, and efficiency (up to 38 %) extend their use to cover base-load. Many simple GT cycle power plants are operating in the GCC, e.g., 497 MW Ras Abu Fontas (RAF) in Qatar including 6 GT×32 MW, 6×48 MW, and 2×9 MW of GT units. It has also 10×7 MIGD capacity multistage flash (MSF) desalting units with steam supplied from HRSG. The plant was commissioned in 1980. The desalting capacity was reported as 55 MIGD in 2010. The GT combined with ST forms GTCC of higher efficiency than either GT or ST cycle. Another diagram of GT is given in Figure 3b.  Figure 3. a: Siemens SGT5-8000H of 400 MW capacity [3] b: Cutaway diagram of a Westinghouse 501D5A gas turbine [4] Simple GT cycle (Figures 4a-d) consists of a compressor, turbine, and generator usually mounted on single shaft and combustion chamber (cc). When started, the generator is usually operated as a motor to get sufficient rotor speed. Then, the GT is ignited, and power supply to the generator-motor is switched off. The GT accelerates until it reaches its nominal speed, and generator is synchronized and connected to the power grid. The GT is operated at constant speed to keep constant frequency at the generator output. The load changes are compensated by the adjustment of the input fuel flow to the combustor.
Another arrangement (Figures 5a, b) is to put the compressor and portion of the turbine on one shaft while the other part of the turbine and the generator on another shaft. The compressor with the first part of the turbine is called the gas generator (GG); and the GG output is equal to the compressor-consumed power. The other part of the turbine and generator is called the free turbine, which produces the net GT power output and gives more flexibility. Figure 4b shows GE's LM2500 Base aeroderivative gas turbine package that has dual fuel (oil and gas) capability; fast load response; 16-stage axial-flow compressor; annular combustor; two-stage, high-pressure, and single-rotor gas turbine; and highly efficient six-stage power turbine.

Analysis of ideal gas turbine cycle
The simple ideal gas turbine open cycle (see Figure 4c) known as Brayton cycle consists of four processes: 1. Isentropic compression of ambient air (working fluid) from pressure P 1 to P 2 by a compressor, with P 2 /P 1 = rp, called pressure ratio.

2.
Heat transfer to the working fluid by mixing fuel with the compressed air and combusted in the cc from 2 to 3; usually P 2 is assumed equal to P 3 for ideal cycle, i.e., isobaric process.

3.
Isentropic expansion of the working fluid in GT turbine from 3 to 4.
The exhausted hot gas is released from the turbine to the atmosphere, and fresh air is used to start or continue the cycle. In fact it is not a real cycle, but process 4-1 can be considered as an isobaric process of heat rejection to atmosphere. The cycle can be represented on both pressurespecific volume (P-v) and temperature-entropy (T-s) diagrams as shown in Figure 4d.
When the air is considered as ideal gas, the property relations for process 1-2 are This is negative work (work consumed by the compressor) and is equal also to , T 2s and h 2s are the absolute temperature and enthalpy at point 2 if the expansion is isentropic (q 1-2 = 0).
Heat addition process from 2 to 3 in the combustion chamber is considered ideal with no pressure loss (isobaric), P 2 = P 3 . The heat input q in between 2 and 3 is equal to the enthalpy increase: It is noticed here that T 3 is the highest temperature in the cycle and is called the turbine inlet temperature (TIT). The amount of specific heat input per kg of air is also equal to where m f is the mass flow rate of the fuel input and LHV is the fuel low heating value (heat generated per kg of fuel, when the water vapor in the combusted gases is in vapor state.
It is also noticed here that w 2-3 = 0.
The property relations of isentropic expansion process in the turbine can be expressed as , , The turbine isentropic work is expressed by This work can also be expressed by enthalpy change as Since part of the turbine work is used to drive the compressor, the net work output (w net = w t -w c ) is expressed as It is noticed that the net heat (q in -q out ) is equal to the net work, or The ideal cycle efficiency (net work/heat in) is expressed by  T  T  P  P  P P P P  T  P  T  P   T  T  T  T  T T  T T  T  T  T  T -- It is noticed that the efficiency depends also on T 3 (TIT), pressure ratio rp, and k, the ratio of specific heats at constant pressure to that at constant volume and is equal to 1.4 for air, The dimensionless work output can be expressed by Process (3)(4): isentropic expansion, w t = c p (T 3 -T 4s ). (5) State (4-1): isobaric heat release, q 4,1 = c p ⋅ (T 1 − T 4s ). (9) There are differences between the ideal Brayton cycle and real gas turbine cycle. In the real cycle, the following are included:

The GT performance
The simple cycle in an h-s diagram including losses is shown in Figure 6a. Figure 6. a: Enthalpy-entropy (h-s) diagram for ideal and practical gas turbine cycle [7] b: Dependence of the thermal efficiency ηth of the cycle on the parameters rp, k, and θ for η t,is = 0.88 and η c,is = 0.86. Line 1 joins points of maximum efficiency for each curve [7] c: Dependence of the specific work of the cycle on the parameters π, κ, and θ for η T =0.88 and η C =0.86. Line 1 joins points of maximum specific work for each curve [7] d: Thermal efficiency vs specific power for varying pressure ratios (10)(11)(12)(13)(14)(15)(16)(17)(18)(19)(20)(21)(22)(23)(24)(25)(26) and combustor outlet temperature (1,473-1,773 K) for a gas turbine [8] ( The losses in turbines are usually expressed by the turbine efficiency defined by η t = actual work of expansion/ideal work of expansion.
Assuming that θ is the ratio of the turbine inlet temperature and compressor inlet temperature, which in this case is θ = T 3 /T 1 , The efficiency of the thermodynamic cycle depends mainly on the TIT (T 3 ) or its dimensionless parameter θ= T 3 /T 1 as well as the pressure ratio rp as shown in Figures 6b-d. The highest cycle temperature is limited by the material and cooling of the first turbine stages; pressure ratio can be optimized to maximize the efficiency for a specific combustor temperature. Besides optimization of the efficiency, the gas turbine is also optimized for power output (Figure 6d). The optimization sets the conditions for the combustor. For the gas turbine cycle in Figure 6 at a combustor outlet temperature at 1,743 K, the optimal pressure ratio for specific power is 14:3 bar and the optimal pressure ratio for efficiency is 25:1 bar. These values are engine specific but show the tendency for optimization. The efficiency of the thermodynamic cycle depends mainly on the TIT (T 3 ) or its dimensionless parameter θ= T 3 /T 1 as well as the pressure ratio rp.
The thermal efficiency always increases with the increase of θ or the TIT, T 3 , which has limitation with the materials. The pressure ratio rp (P 2 /P 1 =P 3 /P 4 ) affects the cycle efficiency, which increases with rp until it reaches a maximum and then starts to fall. The optimal compression ratio changes with alteration of the compressor and turbine efficiencies.
The specific work, defined by the work per unit mass of the air, increases T 3 and reaches a maximum for a certain rp as shown in Figure 6c.
Two distinct losses occur in the combustion chamber: combustion inefficiency and pressure loss.
The first implies an imperfect conversion of the chemical energy in the fuel/air mixture into thermal energy. It is defined as The typical combustion efficiency is around 0.99 or better.
The thermal efficiency of a real gas turbine cycle is lower than the one of the ideal cycle. In the T-s diagram or in the P-v diagram, respectively, the differences are obvious since there are no more isentropic changes possible.

Gas Turbine (GT) components
GTs are operating according to Brayton cycle and using the following components.

Air intake
The air to compressor should pass through an air filter to prevent dust from entering the machine and is accelerated in a duct to the compressor. The inlet duct in front of the compressor is usually designed as a diffuser. This decelerates the air at the inlet and converts part of the air kinetic energy into pressure. Figure 7a shows an air filter installed at the air inlet to the compressor. The inlet air duct can contain an air cooling system. The compression in the GT is a constant volume process. So, the air temperature decrease would increase the air density and mass flow rate, decrease the specific power consumed by the compressor (per unit mass), and increase the GT power output. Figure 7b shows the effect of compressor inlet temperature on the GT output power and heat rate. The air inlet temperature can be decreased by evaporative cooling, fogging, and chilled water system as shown on the psychometric chart given in Figure 7c. Figure 8a shows an inlet air to compressor using evaporative cooling which used relative humidity and wet bulb temperature that are rather low. This system has the advantage of low capital and operation cost as it can operate on raw water and uses air washer that cleans the inlet air. Figures 8b-d show an inlet air to compressor using fogging system. It is also an evaporative cooling system that is used when relative humidity and wet bulb temperature are rather low. This system uses demineralized water and increases GT performance better than the previous evaporative cooling system. Figure 9a shows mechanical refrigeration system (direct type) used hot in areas and can bring the air temperature to any specific requirement irrespective of ambient temperature and humidity ratio. This system has the advantage of increasing the GT performance better than evaporative cooling and fog system. However, this system has high initial capital cost and high operation and capital cost. Figure 9b shows the absorption refrigeration system (direct type), which is similar to that of Figure 9a, but with absorption cooling system operated mainly with steam or hot water substituting the mechanical refrigeration system. This systems has also the advantage of increasing the GT performance better than evaporative cooling and fog system, but at higher initial capital cost and high operation and capital cost.

GT compressor
The main parameters of a compressor are the required pressure ratio (rp), volumetric flow rate, consumed power, and permissible shaft length. The used compressors types in GT application are axial, centrifugal, and combination of both. Axial compressors have more stages to reach the same compression ratio achieved by centrifugal type, and thus, axial compressors have a longer shaft than centrifugal ones. Axial compressors have lower changes of flow direction during compression and thus better efficiency (82-90 %) compared to centrifugal (72-82 %). Axial compressors handle much wider range of volume flows, are used in all heavy utility gas turbines, have much lower tendency for flow separation at the inlet blades, and are more reliable in the case of fast load changes. Centrifugal compressors have small-size, short shafts, used only in small gas turbines (less than 5 MW) and high rotor speeds. Combination of axial and centrifugal compressors utilizes axial compressor reliability and the centrifugal compressor high-pressure ratio.
In centrifugal compressor ( Figure 10a) the air (to be compressed) enters the impeller center and moves outward by centrifugal force to the compressor discharge diffuser. The rotating impellers accelerate the air velocity, and the air kinetic energy is converted to an increase in static pressure by slowing the flow through a diffuser before being discharged.  Figure 8. a: An inlet air to compressor using evaporative cooling which used relative humidity and wet bulb temperature that are rather low [11] b: An inlet air to compressor using fogging system which used relative humidity and wet bulb temperature that are rather low and using demineralized water [11] c: Fog system produces billions of microfine (10-micron average) droplets at 2,000 psi that create a much larger overall evaporative surface, which allows the droplets to evaporate and cool the airflow far more quickly than larger, heavier droplets. This results in faster, more effective evaporation and cooling with significantly lower drain water rates [12] d: MeeFog™ array for a frame 7FA gas turbine, Mee Industries -Fogging Systems for Offshore Gas Turbines (d) Figure 9. a: Mechanical refrigeration system (direct type) used in areas where relative humidity is rather high [1] b: Absorption refrigeration system (direct type) used in areas where relative humidity is rather high [1] Axial compressors have moving (rotor) and fixed (stator) blades ( Figure 8b). The arrays of blades are set in rows, usually as pairs: one rotating and one stationary. While rotating airfoils (known as blades or rotors) accelerate the fluid, the stationary airfoils (known as stators or vanes) decelerate the air, i.e., slow it down, and its kinetic energy is converted to pressure energy. The stators redirect the flow direction for the rotor blades of the next stage. The discharge velocity is almost equal to the suction velocity. This process is repeated by several stages depending on the desired output pressure.  The direction of flow is parallel to the direction of the rotation. The design of compressor blades is different than those of turbines. The compressor blades have divergent profile and act as diffuser to increase air pressure. The turbine blades have convergent profile which works as a nozzle, reducing air pressure by changing its pressure energy into kinetic energy. More on axial compressor design is given in Ref. [15]. Although an axial stage may not offer as much of pressure ratio as a centrifugal stage of the same diameter, a multistage axial compressor offers far higher pressure ratio (and therefore mass flow rates and resultant power) than a centrifugal design.
Separation of the air flow from the surface of the blades of the first compressor stage is real problem in axial and centrifugal compressors. Flow separation from the surface of single blades generates high turbulence in the grid and can partly block the flow path of the incoming air aerodynamically. This effect, called a rotating stall, stresses the whole gas turbine structure with oscillating pressure waves.

GT combustor
The compressed air leaving the compressor is directed to the combustion chamber (cc), called combustor, where fuel such as natural gas (or petroleum liquids) is injected. In a combustor ( Figure 11a) the fuel chemical energy is converted to thermal energy. So, the combustor combines and mixes air and fuel, ignites them, and contains the mixture during combustion. The combustor contains basically four zones -primary zone, secondary zone, dilution zone, and various wall jets -to manage heat transfer at the combustor boundary as shown in Figure  11b. Air entering the combustor is distributed to four major injection points. The first is through swirl vanes positioned at the combustor front face and typically surround the fuel injection port. The swirl vanes impact a circumferential velocity component to the air and thereby thrust the air radially outward as the air enters the combustor (Figure 11c). This creates a pressure void at the center line and induces a backflow to fill the centerline pressure deficit. This effectively creates, as a result, a recirculation flow that extends approximately one duct diameter downstream and defines the "primary zone" of the combustor. The combustors are classified as: 1. Annular (continuous chamber that encircles the air in a plane perpendicular to the air flow) (Figure 12a) 2. Can-annular (similar to the annular but incorporates several can-shaped combustion chambers rather than a single continuous chamber) (Figures 12b, c) 3. Silo (silo, frame-type, combustor has one combustion chamber mounted externally to the gas turbine body) Figure 12. a: Annular combustion chambers [17] b: Gas turbine combustor arrangement [5] c: Several combustors arranged equidistant on the same pitched circle diameter, and each consists of an inner flame tube or liner cylinder mounted on the same axis inside an outer casing cylinder, called tubular combustors [18] Cogeneration Power-Desalting Plants Using Gas Turbine Combined Cycle http://dx.doi.org/10.5772/60209 The combustion process in the GT combustor can be classified as diffusion flame combustion or lean-premix staged combustion. In the diffusion flame combustion, the fuel/air mixing and combustion take place simultaneously in the primary combustion zone, and this generates regions of near-stoichiometric fuel/air mixtures where temperatures and NO x generation are very high. In lean-premix combustion, fuel and air are thoroughly mixed in an initial stage resulting in a uniform, lean, unburned fuel/air mixture which is delivered to a secondary stage where the combustion reaction occurs [19]. The combustion process starts with mixing the fuel with air supported by natural or forced turbulences in the airflow through the combustor. Continuous and stabilized combustion process is affected by the speed of fuel and air particles to the reaction zone, transport of flue gas from there, the speed of the chemical reaction in the reaction zone, and the residence time of any particle in the reaction zone. When the air-fuel mixing is slow compared to the chemical reaction rates, the mixing time controls the burning rate.
In diffusion flames, fuel and oxygen are mixed in the reaction zone through molecular and turbulent diffusion and have wide stability rate of combustion process. It has the advantages of relatively simple design of the fuel nozzles. Since the local conditions at the flame front are rich in fuel, diffusion combustion is insensitive against combustion instabilities and keeps on burning and generates regions of near-stoichiometric fuel-air mixtures with very high temperatures even at very lean conditions. The high temperature by diffusion flames leads to the production of large quantities of thermal NO x .
To reduce the reaction temperatures and/or the formation of thermal NO x , premix combustion is developed, where fuel and air are homogeneously mixed in an initial stage to become lean, unburned fuel-air mixture which is delivered to a secondary stage where the combustion reaction takes place. Manufacturers use different types of fuel-air staging, including fuel staging, air staging, or both; however, the same staged, lean-premix principle is applied. Gas turbines using staged combustion are also referred to as Dry Low NOX combustors. The majority of GT currently manufactured are lean-premix staged combustion turbines.  Figure 13. a: Schematic of a conventional and a lean-premix combustor [20] b: Primary zone temperature influence on NOX and CO emissions [8] In premix, mixing of fuel and air occurs far before the reaction zone. Depending on the burner design and the flow velocity, the time from the fuel injection to the moment of ignition is within several milliseconds. This time is used to create a mostly homogeneous mixture, with a fuel concentration within the ignition range of the specific fuel for the given compressor discharge temperature. The typical adiabatic flame temperature, to which a premix combustion system is adjusted, is at 1,750 K. At this temperature, the formation of NO x is still on an acceptable level, while the heat transfer from the flame is high enough to ensure the ignition of the fresh mixture (Figures 13a, b).
In general, there is an operation window for low emissions that range from the primary zone temperatures 1,670 K to 1,900 K ( Figure 13b). The upper temperature limit is set by the temperature dependence of NOX and the lower limit by carbon monoxide. The increase in CO for lower temperatures is related to poor combustion and the lean blowout limit for the burner.  Figure 14. a: A triple-stage turbine with single-shaft system [17] b: The gas turbine section of the Siemens V94.2 gas turbine. c: Turbine stage with stators to the left which have the main function to act as nozzles to increase the velocity of the gas primarily in the tangential direction, by converting pressure energy to kinetic energy. To the right of the stators are the rotors, which have the function to convert the kinetic energy to power by causing a rotation of the shaft [4] d: Temperature and pressure throughout gas turbine [18]

GT turbine
The hot gases produced in the combustor are expanded in the turbine (Figures 14a-d) to give mechanical energy that operates the compressor, and the balance produces the electric power (EP). The turbine, similar to the compressor, can be axial or centrifugal type. The axial type is easier to cool, as the turbine is exposed to high thermal stresses by the hot gases entering the turbine. The turbine cooling is crucial as it provides the potential of raising the TIT and thus the efficiency. Gas turbines can be particularly efficient when heat content of the hot gases from the turbine is recovered in HRSG to power a conventional ST in GTCC. The hot gases from the GT can also be used for space or water heating or drive an absorption chiller for cooling the inlet air and increase the power output. Figure 14d shows that the hot gases leaving the GT are high enough to generate steam.

The GTCC overview
The exhaust gases leaving the GT can have high temperature (up to 600 ο C) and use a heat recovery steam generator (HRSG) to generate steam. This steam can operate thermally driven desalting units such as multistage flash (MSF) (Figures 15a, b)) or multi-effect thermal vapor compression (ME-TVC) desalting systems or can operate steam turbine (ST). Combination of GT, HRSG, and ST cycle forms GTCC (Figure 15c) of much higher efficiency than single-cycle PP using GT or ST. A schematic diagram of steam turbine (Rankine) cycle components that can be combined with GT is shown in Figure 16a. Large steam turbine is usually divided into high-pressure (HP), intermediate-pressure (IP), and low-pressure (LP) cylinders ( Figure 16b). In GTCC, the GT cycle is called the upper cycle, the steam turbine is called the bottom cycle, and both cycles are shown on T-s diagram in Figure 17a. Modern ST power generation, as shown in Figure 16a, is based on the Rankine cycle which includes the ideal basic cycle processes of (a) isentropic expansion in the steam turbine (ST) from 3-4 and from 5-6; (b) condensation of the steam discharged from the ST in the condenser from 6-1; (c) reversible adiabatic pumping process of condensate from condensing to the HRSG pressures, 1-2; and (d) heat addition at constant pressure in the steam generator (SG) to raise feedwater to saturation temperature, evaporate it, and superheat it from 2-3. In reheat steam cycle, the steam leaving the HP section returns to the SG from 4-5 for further heating before being admitted to the IP cylinder. Reheat is sometimes necessary to raise the steam dryness fraction at the turbine exit than the minimum of 0.88 required by the industry to avoid the blades pitting and raise the efficiency of the LP cylinder.
The use of GTCC to produce both EP gives high-energy utilization factor (UF), up to 80 %, where ( ) UF = Work output + process heat /fuel heat supplied The GTCC is usually used for baseload operations because of its high efficiency. The HRSG can have single-, double-, or triple-pressure stages. The HRSG of single and double-pressure stages and their temperature distribution are shown in Figure 17b. A bottoming steam cycle using double-pressure steam HRSG is shown in Figure 18a. A triple-pressure stage HRSG is shown in Figure 18b. Several differences exist between the steam PP cycle using conventional steam generator (SG) (Figure 19a) and steam cycle in the GTCC (Figure 19b) using HRSG of the GT. The ST plant in Figure 19a has 300 MW electric power (EP) output capacity, using reheat cycle where steam leaving the HP cylinder is reheated in the SG before its introduction to the IP cylinder. This cycle has five closed feed heaters and one open feed heater (deaerator), and the steam flow rate leaving the condenser is 197.86 kg/s, about 76 % that of throttling condition (261.1 kg/s). The ST cycle shown in Figure 19b and with data given in Table 3 is also a reheat cycle of 275 MW output capacity. It utilizes the hot gases leaving three GTs of 164 MW of EP output each. Contrary to the cycle in Figure 19a, the cycle using the HRSG has no feed heaters as all feedwater heating is done in the HRSG, and thus, the steam flow rate leaving the condenser is 438. 8  Net plant efficiency, % (LHV) 44.8 Table 3. Data of the GTCC given in Figure 19b [5]

Steam turbines in GTCC
The steam turbine in the GTCC can be extraction-condensing steam turbine (ECST) ( Figure  20a) or back pressure steam turbine (BPST) (Figure 20b). In the ECST, steam is expanded from inlet pressure (say at 100 bar) and high temperature (up to 538 ο C) to the condenser pressure (about 10 kPa) below atmospheric pressure. As steam expands, its pressure and temperature decrease, while its specific volume and its volumetric flow rate increase. This requires increasing the blade length of the turbine as steam expands to accommodate the increased volumetric steam flow ( Figure 20c). In large-scale steam turbines, the steam volumetric flow is limited by the size of the turbine last stages (see Figure 16b), and this can enforce the use of double-flow condensing steam turbine where the last stage flow is divided between two rows of blades.
In BPST, the steam exits the turbine at the pressure required by the process to be heated as desalination, say 2-3 bar and is higher than that in the end condenser of the ECST cycle, say at 10 kPa. Condensation of discharged steam in industrial processes provides process heat needed for desalination, heating, absorption cooling, or any other processes.
The steam expansion in the ST is usually represented on the enthalpy-entropy (called Mollier chart) as turbine line shown in Figure 21a. For an adiabatic process, the change in enthalpy Δh is equal to the specific work, w per kg of flowing steam. The steam line on the h-s diagram would be a vertical line in reversible (ideal) expansion. The entropy increases during expansion in actual adiabatic process on a Mollier chart. The end point of the irreversible process still lies on that constant-pressure line corresponding to the exhaust pressure. Figure 21a shows that an increase in entropy during expansion decreases the work output, since the change Δh(actual) is less than Δh(isentropic) as the isentropic efficiency defined by: η(isentropic) = Δh(actual)/ Δh(isentropic) <1. One of the main concerns in the design of the ST is its exhaust size selection discharging to the condenser. Lowering the condenser pressure allows more expansion of the steam in the ST, i.e., more decrease in the enthalpy ∆h that is transferred to work. However, decreasing the pressure increases the steam specific volume, thus increasing the steam velocity and increasing the kinetic energy loss of the steam as it leaves the turbine to condenser at almost zero velocity. Figure 21a shows that for the turbine line ABC on the h-s diagram, the exit steam dryness fraction is about 0.84, which is less than 0.88 and not acceptable. Once reheating is done, line ED, the dryness fraction increases to 0.92, which is acceptable. Figure 21b illustrates the exhaust loss curve for a condensing steam turbine. The exhaust area for a particular application should provide a balance between exhaust loss and capital investment in turbine equipment.
Some of the GTCC mount the GT and ST on the same shaft (Figures 22a, b). Since the steam turbine comes to operation after heating up the whole steam cycle, a freewheel clutch is installed between the steam turbine and the generator to prevent the GT from spinning up the steam turbine in a cold steam cycle. Due to the freewheel clutch, the shafts of the gas turbine and the steam turbine are spinning up separately, which prevents them from reaching speed ranges that would cause dangerous resonance frequencies. As soon as the boiler is heated up to operation temperature, the control valve is opened and the steam turbine provides its part of power to drive the generator [23].   Figure 22. a: S107H and S109H single-shaft steam turbine and GT (STAG) equipment configuration [26] b: STAG 107H/ 109H cycle diagram [26] (a) (c) (b) Figure 21. a: Enthalpy-entropy diagram for a steam turbine [25] b: Illustrative exhaust loss curve [25] Cogeneration Power-Desalting Plants Using Gas Turbine Combined Cycle http://dx.doi.org/10.5772/60209

Cogeneration steam turbine
Steam can be extracted from ST for processing heat by using a nonautomatic extraction ST that has openings in the turbine casing for steam extraction, with no means for controlling the pressure of the extracted steam. Steam can also be extracted from an automatic extraction steam turbine with openings in the turbine casing for extraction and means for directly regulating the steam flow to the next turbine stages after extraction opening. Automatic extraction turbines are used when there is a need for process steam at specific pressure between turbine inlet and outlet pressures, as in the case of desalination. There is simultaneous control of the desired extraction steam pressure and turbine speed, even though the demand for extraction steam and the power requirements of the driven load may vary over a wide range. Also an induction-extraction ST that can admit and exhaust steam. In extraction condensing steam turbine (ECST), the steam or part of it exits the turbine at a given pressure and may further be used. The 300 MW steam turbine operating in Kuwait provides full steam demand to two MSF desalting units of 7.2 MIGD each when the turbine EP load varies between 300 and 75 MW.
In Kuwait CPDP, the MSF unit gain ratio defined by desalted water (DW) output to heating supply S (i.e., D/S) has a typical value of 8, and the steam pressure at extraction point to the MSF at full load is 3.5 bar and is throttled to the pressure required by the MSF of 2 bar. When the turbine load is lowered, the steam pressure throughout the turbine is also lowered and reaches about 2 bar at the MSF extraction point when the turbine load is 25 % of the 300 MW nominal load. So, a throttling valve between the extraction point and the MSF is installed to keep the pressure to the MSF plant at 2 bar (Figure 19a). If the steam at the extraction point is less than 2 bar, extraction to the MSF is stopped. In this case, if the MSF can work directly from the high pressure steam supply to the turbine after being throttled and desuperheated (Figures 23a, b).
Steam condensation in the DP provides the steam latent heat as the heating source to the DP. The specific work produced by expanding steam from throttling condition of P 1 and T 1 to the condenser pressure P o and T o is represented by the area encircled by ABCDA in Figure 24a and the area BEFC, represents the specific rejected heat. When steam is extracted at P 3 to the DP, the specific work per kg of steam is represented by AGHD in Figure 23b, and the area GBCH is the work loss for each kg extracted to the DP. In the Kuwaiti plant, the steam to the MSF unit is extracted from crossover pipe between the intermediate-pressure (IP) and the LP cylinders ( Figure 23b). So, the ratio of power to water outputs in CPDP varies as the EP load is always variable and cannot be stored, while water depends on the demand and available storage capacity. So, the EP and DW production ratio is not always constant or matched together.  Figure 23. a: Case 1. HP-LP cylinder features for steam extraction from turbine casing [28] b: Case 2. HP-LP cylinder features for steam extraction from crossover pipe between IP and LP cylinders [28] Steam condensation in the DP provides the steam latent heat as the heating source to the DP. The specific work produced by expanding steam from throttling condition of P 1 and T 1 to the condenser pressure P o and T o is represented by the area encircled by ABCDA in Figure 24a and the area BEFC, the rejected heat. When steam is extracted at P 3 as in Figure 24b to the DP, the specific work per kg of steam is represented by AGHD in Figure 24b, and the area GBCH is the work loss for each kg extracted to the DP.
It is noticed here that in BPST, the steam flow to the turbine depends on the turbine load, and thus, the steam discharged to the DP is slave to the turbine load. So, BPST is usually used in baseload operation, and steam to the DP can be supplied from HP steam line, which is very expensive.

Heat Recovery System Generator (HRSG)
The HRSGs utilize the hot gases leaving the GT to generate steam that can be used to operate thermally driven desalting plants or steam turbines bottoming power cycle. The HRSG can be unfired, supplementary fired or called post-fired (PF), and fully fired. The HRSG can be horizontal or vertical (Figures 25a-d). As given before, the HRSG can have single-, dual-, or triple-pressure level type. The single-pressure stage HRSG has low efficiency, compared to dual-pressure HRSG. In single-pressure HRSG, high efficiency is attained by lowering the stack temperature, and this requires lowering the steam pressure. Lowering the steam pressure lowers the steam cycle efficiency. In dual-pressure designs, lowering stack temperatures would only decrease the first (low)-stage pressure while leaving the second state conditions approximately unchanged. A design parameter of the HRSG is the pinch point (pp), which is the temperature difference between the gas leaving the boiling section and generated steam saturation (or boiling) temperature. The choice of high pp increases the mean temperature difference between the hot gases and water and reduces the heat transfer area but decreases to a certain extent the HRSG efficiency. The low-pressure (LP) generated steam in dual-pressure HRSG can feed the steam turbine at a suitable point or it may be used as process steam for industrial applications (drying, desalination, absorption refrigeration, etc.).
In CPDP, electricity and process heat for desalination are simultaneously produced regardless of gas turbine load; supplementary firing or post-firing (PF) is usually used. In Ras Laffan B CPDP, a very flexible plant design was developed with PF to allow very high thermal power input (maximum 280 MWth) to cope with a wide operational range of GT electrical power and steam production for electricity or desalinated water production.

Cost allocation in CPDP utilizing GTCC
This section develops a mathematical model to evaluate the performance of a typical CPDP using typical GTCC plant and how this performance is affected by parameters such as ambient temperature, compression ratio, air-to-fuel ratio, turbine inlet temperature, and stack temperature. The fuel consumed by the GT is allocated to each of the products (EP and DW) on the basis of the first and second laws of thermodynamics [30]. Figure 26 shows a schematic diagram of a typical CPDP using GTCC, which is considered as reference plant considered here. The plant's design is based on the data given in Table 4 and 50 o C ambient summer temperature and 600 o C temperature of exhaust gases leaving the GT.

Energy analysis
An energy analysis, based on the first law of thermodynamics, is given as follows.

Gas Turbine (GT) cycle
The GT cycle data give 625 o C exhaust gases exit temperature, 50 o C ambient temperature, and 215.5 MW power output for each GT. The isentropic efficiency is 0.85 for the compressor and 0.9 for the turbine. The mechanical efficiency is 0.998 for the turbine and 0.995 for the compressor.

Heat Recovery Steam Generator (HRSG)
There are three GTs, three HRSGs, and only one ST. One third (1/3) of feedwater from the steam cycle returns to each HRSG, and the heat gained by this feedwater is equal to that lost by the exhaust gases, then, where T 4 is the exhausted gas temperatures at the GT exit; T stack is the HRSG stack exit; C p is the gases' specific heat (∼1.11 kJ/kg ο C); h s and h f are the specific enthalpies in kJ/kg of superheated steam leaving the HRSG and feedwater entering the HRSG, respectively; and m s is the steam flow rate from each HRSG. The temperature profile of the hot gases and steam-water temperature in the HRSG is shown in Figure 27. The feedwater is heated from its inlet feed temperature to saturation liquid temperature T sat , evaporated to saturated steam, and then superheated to T s . It is more rational to express the heat supplied to the MSF by its real value in terms of mechanical equivalent energy. The turbine work loss due to discharging its steam to brine heater of MSF unit and not expanding to an end condenser can be calculated; if this steam was expanded in low-pressure (LP) turbine to condenser pressure at 10 kPa and dryness fraction of 0.9, its enthalpy would be 2,345.5 kJ/kg, and the produced work is where Q f,t is the only heat added to the three GT cycles.
where Te is the ambient temperature.
The balance will be unaccounted energy losses,  The high stack temperature (183 °C) is due to the high feedwater temperature returning back from the MSF desalting units at 142 °C.
If GTCC is chosen without desalting plant, lower feedwater temperature is chosen and the stack temperature when NG without sulfur content was used; T stack could be in the order of 100 °C.

Exergy analysis
An exergy analysis, based on the second law of thermodynamics, is conducted here for the cases considered before.

Compressor
The exergy destruction (irreversibility) in the compressor can be presented as follows: where A comp is the increase of flow availability in the air stream across the compressor and equal to ( ) The second law efficiency of the compressor is expressed as

Combustion Chamber (cc)
The main exergy loss (or destruction) of the GT cycle occurs in the combustion chamber (cc) of the GT cycle. An exergy balance in the combustion chamber gives E f = E 3

-E 2 + I cc
The E f , E 2 , E 3 , and i cc are the exergies of fuel input, compressed air inlet, combusted gas exit, and exergy destructed (irreversibility), respectively. This is much lower than the combustion chamber energy efficiency, η cc , based on the thermodynamics first law and is usually assumed equal to 0.99.

Gas turbine
An exergy balance around the GT cycle gives When GTCC is used, the exergy of the exhaust gases E 4 is utilized to generate steam and operate steam turbine, and ε GT in a GTCC is

Heat Recovery Steam Generator (HRSG)
In heat recovery steam generator (HRSG), the heat of hot gases leaving the GT is transferred to feedwater in deaerator, economizer, evaporator, and superheater and the heat transfer in the HRSG.
An exergy balance around the HRSG gives where ΔEg is the exergy loss by the hot gases which is equal to the exergy gain by the water ΔEw plus exergy destruction in one HRSG, I HRSG Then, I HRSG = (ΔE g − ΔE w ) = 12.67 MW , and the effectiveness of HRSG is ε HRSG = ΔE w ΔE g = 0.9 So, the exergy difference gained by water in 3 HRSG = 3 ×116.438 = 349.314 MW, and this is the exergy input to the steam cycle including the three MSF units.

Steam turbine cycle
The exergy difference across the ST cycle, ⊗EST, is equal to (Esi -Ese), where Esi and Ese are the exergy of the steam inlet to the turbine and steam outlet to the MSF units, respectively.

Desalination system
The exergy difference between the discharged steams from the turbine to the condensate from the brine heaters of one MSF unit, ∆Ede is as follows: where m sd and (h d,in -h d,e ), s d,in , and s de are the steam flow rate to each MSF desalting unit, its specific enthalpy difference between the steam inlet, and its condensate exit from the desalting unit, specific entropy at steam inlet, and specific entropy of its condensate at the exit, respectively.
So, an exergy balance of the given CPDP using GTCC and MSF units shows that there are unaccounted losses due to steam extracted at moderately high pressure to operate the steam ejectors of the MSF units, and others and can be calculated as follows:

Exergy distribution of the overall GTCC
Exergy balance of the given CPDP using GTCC and MSF units is conducted for the whole GTCC cycle.
The fuel exergy of the fuel supplied for the three GTs is There are unaccounted losses due to steam extracted at moderately high pressure to operate the steam ejectors of the MSF units, and others and can be calculated as follows: .

Fuel allocation between the EP and DW production
There are two methods to allocate the fuel between the EP and DW production, while the second is exergy method.

Work loss method
The first method is the work loss method. As mentioned, there is work loss due to discharging steam to the MSF units instead of its expansion to the condensing turbine. It is clear that both methods give very close results, but the first method is easier and understandable by practitioner engineers.

Desalinated water cost
Since all combined cycle power plants in Kuwait were dual fuel (i.e., can be operated either by natural gas or heavy oil), the cost of desalinated water is evaluated in this section based on the current oil and natural gas prices. Hence, the desalinated water produced by this plant will be estimated based on two different types of fuel as follows: The oil price is 60 $/bbl and the low heating value of the oil is LHV oil = 42229 kJ/kg; so, the energy content in 1 barrel of oil (density of 900 kg/m 3 ) can be calculated as follows: 1 barrel = 0.159 m 3 × 900 kg m 3 × 42299 kJ kg = 6. 04 GJ; so, the oil price per GJ will be $9.33/GJ The following assumption is assumed for the analysis: steady state operation.

Sensitivity analysis
The developed equations were used to evaluate the performance of the reference CPDP using the GTCC in this section. A simplified schematic diagram of plant is shown in Figure 30, while state point conditions of the model are given in Table 5.
The model was tested against the available data of Al-Shuaiba CCPP, and the results showed good agreements as shown in Table 4. This model can also be used for simulation and/or parametric studies of the plants in order to evaluate its performance. A sensitivity analysis is carried out to investigate the effects of some combined cycle parameters on the overall efficiency, specific fuel energy to desalination, as well as the desalinated water cost. The selected parameters are ambient air temperature, compression ratio, and air-to-fuel ratio, turbine inlet temperature, and stuck temperature.   The effect of ambient air temperature on the fuel allocation between the electric power and desalinated seawater production is presented in Figure 31. It shows that as the ambient air temperature increases, the allocated fuel to the electric power decreases, and this led to increase the allocated fuel to desalination. Figure 32 shows the effect of ambient temperature on the combined cycle efficiency at different compression ratios. It is clear that the cycle efficiency is the highest at maximum pressure ratio and minimum ambient temperature. On the other hand, the effect of air-to-fuel ratio is limited as shown in Figure 33. Figure 34 shows the effect of ambient and turbine inlet temperatures (TIT) on the specific fuel energy to desalination. It shows that the specific fuel energy to desalination increases at high ambient temperatures, while it decreases at higher turbine inlet temperatures.       On the other hand, the effect of stack temperatures on the heat rejection into environm destruction of HRSG is depicted in Figure 35. It is clear that as the stack temperatures into the environment will increase and consequently the exergy destruction in HRSG wi Although the effect of steam turbine inlet pressure on the specific fuel energy to des slightly be affected by the types of fuel as shown in Figure 37. On the other hand, the effect of stack temperatures on the heat rejection into environmen destruction of HRSG is depicted in Figure 35. It is clear that as the stack temperatures i into the environment will increase and consequently the exergy destruction in HRSG will d Although the effect of steam turbine inlet pressure on the specific fuel energy to desali slightly be affected by the types of fuel as shown in Figure 37. On the other hand, the effect of stack temperatures on the heat rejection into environment as well as on the exergy destruction of HRSG is depicted in Figure 35. It is clear that as the stack temperatures increases, the heat rejection into the environment will increase and consequently the exergy destruction in HRSG will decrease. Although the effect of steam turbine inlet pressure on the specific fuel energy to desalination is negligible, it will slightly be affected by the types of fuel as shown in Figure 37.  Figure 38 shows the effect of ambient temperature on the specific fuel energy to desalination when different types of fuel are used. Since there is a large difference in price per unit of energy between oil and natural gas, the cost of desalinated water is strongly affected by the fuel type as shown in Figure 39.  The effect of steam turbine inlet pressure on the specific fuel energy to desalin Figure 37. The effect of ambient temperature on the specific fuel energy to desalination at different Figure 38 shows the effect of ambient temperature on the specific fuel energy to desali fuel are used. Since there is a large difference in price per unit of energy between o desalinated water is strongly affected by the fuel type as shown in Figure 39.   Figure 38 shows the effect of ambient temperature on the specific fuel energy to desalin fuel are used. Since there is a large difference in price per unit of energy between oi desalinated water is strongly affected by the fuel type as shown in Figure 39.

Conclusion
A general overview on the CPDP using GTCC is presented including d component. Energy and exergy analyses, based on the first and second laws conducted on CPDP using GTCC connected with MSF desalting system. The steam from the ST at higher pressure and temperatures compared to end co calculated. The exergy at different points of both GT and ST cycles and HRSG components were calculated. The main exergy loss was found in the GT c allocation between the desalting process and power production was conduct methods. Both methods gave almost the same results.

Conclusion
A general overview on the CPDP using GTCC is presented including description and analysis of the GTCC component. Energy and exergy analyses, based on the first and second laws of thermodynamics, respectively, were conducted on CPDP using GTCC connected with MSF desalting system. The concept of work loss due to exhausting steam from the ST at higher pressure and temperatures compared to end condenser condition was introduced and calculated. The exergy at different points of both GT and ST cycles and HRSG and the exergy destructions in several components were calculated. The main exergy loss was found in the GT combustion chambers. The fuel energy allocation between the desalting process and power production was conducted, based on the work loss and exergy methods. Both methods gave almost the same results. The main problem detected from the design of the given plant the high stack temperature of 183 o C of the HRSG to match that of high feedwater returning from the MSF desalting units at 135 o C. In GTCC using condensing turbines and NG with no sulfur, the typical HRSG stack temperature is 100 o C. Sensitivity analysis shows that the pressure ratio, inlet air temperature, turbine inlet temperature, and stack temperature have a significant role in the combined cycle performance. It shows also that the cost of desalinated water is strongly affected by the fuel type because there is a large difference in price per unit of energy between oil and natural gas.