Analysis on Solar Retrofit in Combined Cycle Power Plants

Master Thesis

Analysis on Solar Retroﬁt in Combined Cycle Power Plants

ausgefuhrt¨ zum Zwecke der Erlangung des akademischen Grades eines Master of Science unter Anleitung von

Univ.Ass. Dipl.-Ing. Armin STEINER Univ.Prof. Dipl.-Ing. Dr.tech. Markus HAIDER

E 302 - Institut fur¨ Energietechnik und Thermodynamik

erstellt an der

Technische Universit¨at Wien Fakult¨at fur¨ Maschinenwesen und Betriebswissenschaften von

Andrea MIGUEZ DA ROCHA

Wien, am 23.04.2010 Contents

Contents

1 Introduction 1

2 Combined Cycle Power Plants 2 2.1 GasTurbine...... 2 2.1.1 Improvements for increasing the work output ...... 3 2.1.2 Part-load performance ...... 6 2.2 Steam Cycle ...... 9 2.2.1 Types of Condensers ...... 11 2.3 Heat Recovery Steam Generator ...... 13 2.3.1 Types of HRSGs ...... 14 2.3.2 Design considerations ...... 15 2.4 Combined Cycle Power Plants ...... 17

3 Concentrating Solar Power Plants 20 3.1 Solar Thermal Concentrating Collectors ...... 20 3.1.1 Parabolic-Trough Power Plants ...... 20 3.1.2 Dish Stirling Systems ...... 22 3.1.3 Central Receiver System ...... 23 3.2 Parabolic Trough Power Plant Conﬁgurations ...... 25 3.2.1 SolarMode ...... 25 3.2.2 Direct Steam Generation ...... 25 3.2.3 Integrated Solar Combined Cycle (ISCC) ...... 26

4 Modeling and Simulation with EBSILONr Professional 27 4.1 Basics of the Ebsilon Software ...... 27 4.1.1 Data introduction ...... 28 4.1.2 Calculation Modes ...... 29

5 Model of the Gas Turbine GE 9FA 30 5.1 Gas Turbine GE 9FA in Design Conditions ...... 30 5.2 Gas Turbine GE 9FA in Off-design ...... 32 5.2.1 Variation of Ambient Temperature ...... 32 5.2.2 Variation of fuel and air flows ...... 35 5.2.3 Variation of Efficiency in Components ...... 38 5.3 Off-load Model of HRSG, Steam Turbine and Condenser ...... 40

6 Modeling a Single Pressure CCPP 42

7 Modeling a Two Pressure CCPP 46

8 Modeling a Three Pressure CCPP 48 8.1 Performance of Three Pressure CCPP in ISO Conditions ...... 49 8.2 Performance of Three Pressure CCPP with Changing Ambient Temperature 49

ii Contents

9 Standard Three Pressure CCPP with Solar Boosting 55 9.1 Standard Three Pressure CCPP with Solar Boosting and Saturated High Pres- sure Steam (SHP) ...... 56 9.2 Standard Three Pressure CCPP with Solar Boosting and Cold Reheated Steam (CRH) ...... 59

10 Conversion Efficiency of Solar Boosting 68 10.1 Conversion Efficiency of Saturated High Pressure Steam ...... 68 10.2 Conversion Efficiency of Cold Reheated Steam ...... 68

11 Impact of Solar Steam Injection on Temperatures and Pressures 68 11.1 Impact of Saturated High Pressure Steam Injection ...... 68 11.2 Impact of Solar Heat Injection as Cold Reheated Steam ...... 69

12 Thermal Eﬃciency of the CCPP 70 12.1 Thermal Eﬃciency of the CCPP with Solar Boosting ...... 70

13 Standard Three Pressure CCPP with Increased Solar Boosting 70

14 Economic Analysis of the Three Pressure CCPP with Solar Boosting 72 14.1 Pneumatic Pre-Stressed Concentrators (PPC) ...... 75 14.2 Parabolic Trough Concentrators (PTC) ...... 77

15 Economic Analysis of the Three Pressure CCPP with Increased Solar Boosting 78 15.1 Pneumatic Pre-Stressed Concentrators (PPC) ...... 79 15.2 Parabolic Trough Concentrators (PTC) ...... 80

16 Conclusion and Summary 80

Bibliography 82

iii 1 Introduction

1 Introduction

The objective of the thermodynamics studies of thermal power plants is the determina- tion and maximization of the eﬃciency of combined cycle power plants. With other words, we wish to study and apply the methods for increasing the plant eﬀectiveness by saving costs.

In the production of electricity the combined cycle power plants are largely known and developed. A combined cycle power plant is the combination of two cycles associated with the power production, Rankine and Brayton.

The Rankine cycle consists in a close steam cycle where the steam at high pressure is expanded in a turbine to produce power. The Brayton cycle is an open cycle in which air enters in a compressor for being mixed with fuel in a combustor chamber. The mixture of fuel and air, at high pressure and temperature, is expanded through a turbine which produces power. Both cycles are combined in a way that the energy in the exhaust gases of the gas turbine supports the steam cycle.

Gas Turbines are volumetric machines. At high ambient temperature the air and gas ﬂow through the gas turbine decreases due to the lower air density. The gas turbine produces less power. The steam cycle is also over dimensioned due to the fact it is receiving less exhaust energy than in its design.

In this study we propose a solar retroﬁt for making up for the lack of input energy in the steam cycle from the gas turbine. Linear concentrating thermal technology is applied. The solar retroﬁt can deliver additional steam for the steam cycle without spending the investment costs of power island. Therefore the concept can be economically advantageous.

In the following process models of a single, two and three pressure combined cycle power plants will be developed and simulated with the software EBSILONr Professional 8.0. The diﬀerent modes of operation will be analyzed. The performance of the three pressure combined cycle will be modeled under diﬀerent load conditions.

Two diﬀerent modes of boosting the combined cycle power plant will be studied and compared: boosting the steam cycle with saturated steam at high pressure (SHP) or with cold reheated steam (CRH). The chosen solution will be analyzed both technically and economically comparing two concentrating thermal technologies, Parabolic Trough Concentrators (PTC) and Pneumatic Pre-Stressed Concentrators (PPC).

1 2 Combined Cycle Power Plants

2 Combined Cycle Power Plants

The combined cycle is one of the most eﬃcient cycles in operation for power generation. Its thermal eﬃciency can reach 66%. In design conditions, the gas turbine supplies the 60% of the power, while the steam turbine delivers only the 34% of the energy.

The combined cycle exists in different configurations: single-shaft or multi-shaft. The difference is the number of gas turbines and heat recovery steam generators (HRSGs) delivering power to the steam turbine.

2.1 Gas Turbine A gas turbine is an engine which allows the conversion of the energy of fuel in some form of useful power, such as mechanical power. The simplest cycle of a gas turbine is formed by a compressor, where the air is compressed until the required pressure; a combustion chamber, where the fuel and air at high pressure are mixed, and a turbine, where the mixture of gases are expanded after the combustion. In the turbine, the gases are expanded adia- batically generating a big amount of work. Part of the work obtained, between the 50% to 60%, is used to drive the compressor an the rest is delivered to the surroundings, Bathie [2].

The ideal Brayton cycle is shown in Figure2.1. The cycle eﬃciency is given by the following

Figure 2.1: Simple Brayton cycle, [17] relationship: (γ−1) 1 γ η = 1 − (2.1) r and depends only on the pressure ratio (r) and the nature of the gas (γ is the adiabiatic coeﬃcient of the air).

By taking into account in the overall cycle efficiency, the efficiencies of the compressor (ηc) and the turbine (ηt) which work between the firing temperature (Tf ) and the ambient tem-

2 2 Combined Cycle Power Plants

perature (Tamb), we can obtain the following equation:

 γ−1  ( γ ) Tambr ηtTf − 1 η =  ηc  1 − (2.2) cycle  γ−1  γ−1  ( γ )  ( γ ) T − T − T r −1 r f amb amb ηc

The Figure 2.2 shows how increases the cycle eﬃciency when the pressure ratio and the ﬁr-

Figure 2.2: Overall Cycle Efficiency of the Pressure Ratio [3] ing temperature increase. At a given firing temperature, the efficiency increases with higher pressure ratios; however, increasing the pressure ratio too much over a specific value can result in the opposite effect, lowering the overall efficiency.

The optimum pressure ratio for getting the maximum power output of the turbine can be expressed by the following relationship:

γ 2−2γ Tamb 1 ropt = ηt (2.3) Tf ηc

Comparing Figure 2.2 and Figure 2.3, we can come to the conclusion that the pressure ratio for reaching the maximum eﬃciency is much higher than for reaching the maximum work per kg of air.

It is necessary to be noted, that having a look at the eﬃciency equation we can conclude that the overall eﬃciency of the cycle can be improved by increasing the pressure ratio, decreasing the compressor inlet temperature or increasing the turbine inlet temperature.

2.1.1 Improvements for increasing the work output There are many ways of improving the performance of the basic gas turbine and raising the cycle eﬃciency. We will discuss in the following three of the most common possibilities:

3 2 Combined Cycle Power Plants

Figure 2.3: Pressure Ratio for Maximum Work per kg of Air[3]

Regeneration

The regeneration effect takes place by adding a heat exchanger which heats the air at its entrance reducing the fuel consumption. The efficiency can be increased because the flue- gas is hotter in the exhaust of the turbine than the air leaving the compressor. With the regeneration is possible to preheat the air before the combustion chamber so the necessary amount of fuel is smaller. This increases the cycle thermal efficiency without changing the work produced in the cycle, Bathie [2].

It is important to note that the higher effectiveness the regenerator the bigger heat exchanger, because the transferring area has to be also bigger. That means the pressure drop in the heat exchanger and the space required for the installation increase which lead to higher costs. It is necessary to find the equilibrium between increasing efficiency and lowering costs.

Intercooling

When the air enters the compressor a certain amount of work is needed to rise the pressure from the ambient to the turbine pressure. That amount of work can be reduced by dividing the compression process in two or more stages and cooling the air in between. Lowering the temperature of the air lowers the volume too and more mass of air can be compressed at the same time. Besides, dividing the process in more parts permits to approximate the

4 2 Combined Cycle Power Plants

Figure 2.4: Regenerative Brayton cycle [4] compression process to an isothermal process. The problem is with more stages the process becomes more expensive because more compressors are needed.

Figure 2.5: The Intercooled Gas Turbine Cycle [3]

Although the work given for the turbine does not increase, the work required for the compressor is lower. The result is, with intercooling more work can be delivered to the generator because the compression process needs less work.

It is important to say the work increases but does not occur the same with the efficiency of the cycle. Dividing the gas turbine cycle into a lot of cycles (m − n − o − p − m) as in Figure 2.5 it is possible to simulate the Carnot cycle. The efficiency for the Carnot cycle is given by: Tm ηCARNOT = 1 − (2.4) Tp and when the specific heats are constant, we can assume

γ−1 T T T P γ 3 = m = 2 = 2 (2.5) T4 Tp T1 P1 By joining all the Carnot cycles the gas turbine cycle can be simulated. But taking into account the additional part of the cycle due to intercooling (a − b − c − 2 − a) the eﬃciency

5 2 Combined Cycle Power Plants will decrease because in that part the eﬃciencies of the Carnot cycles are smaller.

An intercooling regenerative cycle can increase the power output and the thermal eﬃciency. This combination provides an increase in eﬃciency of 12% and an increase in power output of about 30%, Boyce [3].

Reheat

The reheat is based in the same theory than the intercooling, but in this case the expansion is divided in stages in an intent to aproximate the expansion process to an isothermal process. For reheating, two turbines are necessary and between the ﬁrst and the second stage the products of the combustion are reheated until the maximum temperature of the cycle. That permits to obtain more work during the second part of the expansion process.

Figure 2.6: Reheat cycle and T-s diagram [3]

There is an intermediate pressure which divides the stages for both process, reheat and intercooling. The optimum value for that pressure is given by the following equation, Boyce [4].

√ pi = p1p2 (2.6)

2.1.2 Part-load performance The variation of thermal eﬃciency with reduction in power, sometimes referred to as partload performance, is of major importance in applications where considerable running at low power settings is required, Saravanamuttoo [1].

When determining the off-design performance it is important to be able to predict not only the effect on thermal efficiency at part load, but also the effect of ambient conditions on maximum power output, the effects of high and low ambient temperatures and pressures must be considered. The variation of maximum power with ambient conditions is clearly of prime importance to the customer, and the manufacturer must be prepared to guarantee the performance available at any specified condition. Cold conditions are beneficial for power output and result in a modest increase in efficiency, while warmer conditions cause large

6 2 Combined Cycle Power Plants decrease in both power output and eﬃciency.

The start point for studying the ﬂow rate behaviour of a multistage turbine is the inves- tigation of the behaviour of a single stage. If we apply the continuity equation to a cross section in a nozzle, every unit of section is being crossed for an amount of ﬂow: m˙ = ρc = µρscs (2.7) Amin

where µ is the flow rate, ρ and c the medium magnitudes, and ρs and cs, the isentropic medium values of density and velocity in the cross section. The flow rate depends of the efficiency and because of that depends also of Reynolds and Mach. In most of the practical cases the variation of µ is very small because Reynolds values are high enough.

The velocity in the outlet of the stage is calculated with the total pressure at the inlet, pA, and the static pressure at the outlet, pB: v " κ−1 # u κ u 2κ pB c = t pAvA 1 − (2.8) κ − 1 pA

For the density ρ = 1/v

1 κ 1 pB ρs = (2.9) vA pA

Combining equations 2.8 and 2.9, we obtain:

r s 2 κ+1 p 2κ p κ p κ rp m˙ = µA√ A B − B = µA A ξ (2.10) pAvA κ − 1 pA pA vA

The capacity is described by the next function v " 2 κ+1 # u κ κ u 2κ pB pB ξ = t − (2.11) κ − 1 pA pA

For the sound pressure ratio

κ p 2 κ−1 Bκ = (2.12) pA κ + 1 it is possible to obtain the maximum value

1 r 2 κ−1 2κ ξ = (2.13) max κ + 1 κ + 1

If the pressure ratio decreases below the critical value compare to the equation (2.12), the

7 2 Combined Cycle Power Plants capacity does not change, that means that the capacity function ξ remains at the maximum level. By leaving κ = constant, the relative capacity function: ξ φ = (2.14) ξmax which looks like in Figure (2.7).

Figure 2.7: Relative ﬂow rate function, [15]

The behaviour presented is associated with an ideal single nozzle. When we connect more nozzles in series, we can make an approximate statement for a complete turbine. The flow is in critical conditions in such a configuration, equation (2.12) has to be filled for at least one nozzle. The product of the pressure ratio of all nozzles has to be smaller than in equation (2.12), due to the fact that the pressure ratio, pB/pA, for the rest of nozzles have to be below the unity.

The point K in the Figure 2.8 slices to the left until ﬁnally, for a multi-stage turbine, we obtain an ellipse as the curve which describes the turbine behaviour.

For this case the function can be approximated for a circle, s p 2 φ = 1 − B (2.15) pA

Furthermore, the flow rate is influenced by the rotation of the turbine. Therefore the function µ¯ is formed. Now it is possible to describe the ratio of two flow rates by using equation (2.10), where the index ”0” marks the design case.

m˙ µ¯ p rp v φ = A A0 A0 (2.16) m˙ 0 µ¯0 pA0 pA vA φ0

8 2 Combined Cycle Power Plants

Figure 2.8: Relative ﬂow rate function, [15]

κ−1 Taking into account the caloric equation of state pAva = κ hA, we get: m˙ µ¯ p rh 0 φ = A A (2.17) m˙ 0 µ¯0 pA0 hA φ0

For a multi-stage turbine the inﬂuence of the revolution speed is small, so we can state µ/¯ µ¯0 = 1. Finally, we write the simpliﬁed equation as:

r s 2 2 m˙ hA0 pA − pB = 2 2 (2.18) m˙ 0 hA pA0 − pB0

In steam turbines the conditions at the entrance can be changed by the throttling, so it is possible to say that hA = hA0. If hA = hA0 = constant the equation (2.18) describes a cone cylindrical surface.

In steam turbines and stationary gas turbines the case of constant power output have a practical meaning. With pB = pB0 = constant and a ﬁxed ratio hA/hA0 the equation (2.18) describes a hyperbola. That is described in the Figure 2.9, withm ˙ as abscissas and pA as ordinates. hA/hA0 appears as parameter. For ratios pA/pB > 4 the hyperbola can be quite exactly replaced with its asymptote in the way that the ﬂow ratio will be directly proportional to the initial pressure.

m˙ rh p = A0 A (2.19) m˙ 0 hA pA0

In the turbine condenser the ﬁnal pressure has only a hundredth bar, that means that prac- tically pB = pB0 = 0 can be applied in the (2.19) for all the pressure range.

2.2 Steam Cycle The steam turbine is an engine in which a steam ﬂow, at high pressure and temperature, is expanded transforming its energy into kinetic energy, which is as well converted into work by moving the rotational parts of the turbine, Boyce [3]. The performance of the steam turbine is described by the Rankine cycle.

9 2 Combined Cycle Power Plants

Figure 2.9: Ratio between mass ﬂow and inlet pressure by leaving the outlet pressure constant, [15]

The Rankine cycle is the most common thermodynamic cycle utilized in the production of electrical power with water-steam as the working ﬂuid and consists in two isobaric and two adiabatic processes. Pumped water from low to high pressure enters a boiler where it is heated at constant pressure by an external heat source to become a dry saturated steam. The saturated steam expands through a turbine, generating power and decreasing its temperature. The wet steam enters a condenser to become a saturated liquid.

Figure 2.10: Rankine cycle

10 2 Combined Cycle Power Plants

The thermal eﬃciency of the Rankine cycle is given by the following expression: (W˙ − W˙ ) W˙ η = turb pump ≈ turb (2.20) ˙ ˙ Qin Qin

The work required for pumping the water is very small compared with the work that the turbine produces, so is not taken into account at the time to estimate the eﬃciency.

In a real Rankine cycle, the compression in the pump and the expansion in the turbine are not isentropic. In other words, these processes are non-reversible and entropy is increased during the two processes. This increases the power required by the pump and decreases the power generated by the turbine.

In particular the eﬃciency of the steam turbine will be limited by water droplet forma- tion. As the water condenses, water droplets hit the turbine blades at high speed causing pitting and erosion, gradually decreasing the life of turbine blades and eﬃciency of the turbine. The way used to avoid this problem is superheating the steam.

As for the gas turbine, there are improvements for increasing the eﬃciency of the cycle:

In the reheated cycle two turbines work in series and after the ﬁrst expansion in the high pressure turbine the steam re-enters the boiler and is reheated almost until the maximum temperature of the cycle. Then pass through the second, lower pressure turbine. Among other advantages, this prevents the vapor from condensing during its expansion which can seriously damage the turbine blades, and improves the eﬃciency of the cycle.

In the regenerative Rankine cycle the water after emerging from the condenser (possibly as a subcooled liquid) is heated by steam tapped from the hot side of the cycle.

2.2.1 Types of Condensers In the condenser the steam leaving the turbine in the steam turbine is condensate to water. The steam quality before entering the turbine is usually between 90% to 96% which means 10% to 4% of liquid in the mixture, Boyce [3].

˙ The amount of heat removed (Qc)by the condenser is given by the next equation: ˙ Qc =m ˙ s(hs − hc) (2.21) wherem ˙ s is the mass ﬂow of steam through the condenser, hs is the enthalpy of the steam leaving the steam turbine and hc is the enthalpy of the liquid leaving the condenser.

In condensers, working with water or with air, the amount of heat extracted from the steam has to be the same that the cooling ﬂuid receives.

m˙ s(hs − hc) =m ˙ air(ho − hi) =m ˙ wCpw(To − Ti) (2.22)

11 2 Combined Cycle Power Plants

Figure 2.11: Regenerative and reheated Rankine cycle, [22]

wherem ˙air andm ˙w are respectively the mass of air and water and the the subscripts ”o” and ”i” refers to the values at the condenser output and inlet.

The overall thermal transmittance (U) in a condenser is the amount of heat transmitted per unit of time, unit of surface and degree of temperature difference being the property which defines the behavior of the condenser. The thermal transmittance is written as in the equation (2.23): m˙ C θ U = w pw Ln 1 (2.23) A θ2 where θ1 is the temperature difference between the steam and the cooling water entering the condenser, and θ2 is the temperature difference between the the steam and the cooling water after passing through the condenser.

The most common types of condensers used in combined cycle power plants are the water- cooled condenser and the air-cooled condenser.

Water-cooled Condenser

In this type of condensers, the cooling water is the refrigerant fluid and it removes the heat from the steam flow condensing the steam to water. That condenser consists of a bunch of pipes through which the cooling water flows while the steam flows out of the pipes. The cooling water can be provided from a river or the sea, from where it is pumped to the condenser. Before returning to the river or sea the water is put in a holding pond. The water-cooled condenser can be either classified in shell-and-tube or in tube-and-tube type, Petchers [6].

Air-cooled Condenser

12 2 Combined Cycle Power Plants

The Air-cooled condenser is used in places where it is difficult to find a source of cooling water. This is generally the most expensive type of condenser. When the ambient temperature is high the resulting operation temperature in the condenser can reduce significantly the efficiency of the refrigerator system. Air-cooled condensers do not need water treatment and have less problems of freezing up with cold temperatures. Air-cooled condenser require less maintenance than other types but the service life is shorter than water-cooled condensers due to the coil degradation.

In air-cooled condensers the steam ﬂow is condensed into ﬁnned tubes by ambient air. The cooling air is moved by axial fans, which are moved by electric motors.

Cooling Tower

A cooling tower is a heat exchanger in which two ﬂuids, water and cooling air, are put in direct contact to transfer heat. The process consists of a water ﬂow which is sprayed into a rain-like pattern, through which the ambient air is induced by fans.

2.3 Heat Recovery Steam Generator The heat recovery steam generator (HRSG) is an important subsystem of a combined power plant which uses the energy from the exhaust gases of the turbine for transferring heat to water and generating steam at high temperature and pressure.

The exhaust gases leave the gas turbine at approximately ambient pressure and at very high temperature (500◦C to 600◦C). This energy is used for the HRSG to produce steam. Although there are many conﬁgurations of HRSG, most of them are divided in the same number of sections as the steam turbine. There is one section for high pressure (HP), low pressure (LP) and sometimes for an intermediate pressure (IP). Each section of the HRSG has a evaporator or steam generator and depending on the section is possible to ﬁnd an economizer or a superheater, Boyce [4].

A HRSG is composed basically of individual heat exchangers which exchange the energy from the exhaust gases of the turbine with the water/steam of the Rankine cycle.

The water enters ﬁrst in the economizer for being pre-heated and then goes through the evaporator where the steam is generated at constant pressure and temperature. Finally the steam is superheated in the superheater. After that, the superheated steam enters the turbine where it is expanded and the power is generated.

The majority of the heat is transferred by convection and for increasing the heat surface finned tubes are used. As the heat transfer on the waterside is much higher than on the exhaust gas side, due to the bigger temperature difference between gas and water than between gas and steam, the fins are used on the gas side to rise the heat transfer.

13 2 Combined Cycle Power Plants

2.3.1 Types of HRSGs The selection of the HRSG depends of many factors but the most determinant are the initial cost and the global eﬃciency of the plant. A 1% of increase in eﬃciency leads to a 3-4% of increase in the costs, Boyce [3].

The most common units used in combined cycle power plants are the the drum type HRSG with natural circulation. The type shown in Figure 2.12 is formed by separate components: drums, economizers, superheaters, generating tubes and blowdown systems. The tubes are disposed vertically and the exhaust gas flow is horizontal. In HRSGs with forced circulation, the mixture of steam and water is pumped through the evaporator tubes. But the use of pumps brings to the cycle a parasitic load lowering the efficiency of the cycle. There are some vertical HRSGs with natural circulation or the other possibility are the Once Through Steam Generators, which do not have the defined sections as we can find in the drum HRSG. The OTSG is a pipe in which the water enters at one side and the steam leaves the pipe at the other side. There are no evaporators and no drums so there is no defined point where the evaporation takes place and the interface between water and steam changes its position depending on the heat supplied.

Figure 2.12: Section of an horizontal HRSG, [20]

The three diﬀerent types of HRSGs are shown in Figure 2.13.

Water is being delivered to the drum with the help of the pumps. In the drum, the

14 2 Combined Cycle Power Plants

Figure 2.13: Types of HRSGs, [20] feed water is mixed with the mixture of water/steam being the steam separated due to the higher density of the water.

The circulation ratio can be described as the ratio between the mass ﬂow of water circulating into the evaporator and the mass ﬂow of steam leaving the evaporator, Effenberger [21] .

m U = circ (2.24) ms

The circulation ratio depends on the pressure and normal values for natural circulation are from 5 to 50 and pressures around 160 bar, while for forced circulation normal values are between 3 to 10 and pressures around 180 bar.

Other possible classification for HRSG can be made depending on supplementary firing. There are unfired, supplementary fired and exhaust fired HRSGs. In unfired HRSGs, the energy supplied for the turbine is used without changes, while in the others the mass of gas from the turbine is mixed with additional fuel to increase the steam production, Boyce [3].

One of the advantages of the supplementary ﬁring is that the system is able to follow the demand, providing more power at peak loads. In these systems the gas turbine is sized according to base load demand, but it has to deliver enough power for higher peaks loads.

2.3.2 Design considerations Some other features should be taken into account in the design of HRSGs.

Pinch Point

15 2 Combined Cycle Power Plants

The pinch point is the temperature diﬀerence between the gas at the evaporators exit and the steam saturation temperature. For lowers pinch points the transferred heat is higher. But the heating surfaces have to be bigger with the consequent increase of pressure drop and costs. Higher pinch points could mean a lower production of steam. The range for normal values of pinch points is between 8-22◦ C.

Approach Temperature

The approach temperature is deﬁned as the temperature diﬀerence between water at the evaporator inlet and the steam saturation temperature. The production of steam is higher at small approach temperatures. Typically, approach temperatures are in the 5.5-11◦ C range.

The Figure 2.14 represents the q,T diagram for one section of the HRSG where we can see the evolution of the water and steam going through the HRSG:

Oﬀ-design performance

The gas turbine performance and in particular the exhaust energy determines the operation of the HRSG. The variation in the ambient conditions, the load or the own health of gas turbine change the output conditions of the gas ﬂow, changing the behavior of the HRSG too.

Gas turbines without inlet air temperature cooling deliver at high ambient temperatures less power due to the higher air density, increasing the volume of the air and decreasing the mass ﬂow of air which can be compressed at the same time. As Ganapathy [5] says, the power output could variate as much as 15-25% between the coldest and hottest temperatures. To avoid this change of performance the possibilities are use evaporative coolers, mechanical or absorbtion chillers or thermal storage can be used.

Temperature (◦ C) -6.7 4.4 15.6 26.7 37.8 48.8 Power (kW) 38.15 38.6 35.02 30.82 27.36 24.04 Exhaust Temperature(◦ C) 390 415.6 425 437.8 450.6 465.6 Exhaust Flow (kg/s) 141.52 137.89 129.73 119.75 110.68 102.06

Table 1: GT Performance at diﬀerent ambient temperatures. Data for LM 5000 gas turbine, [5]

In Table 1 and Figure 2.16 we can see that the power output in the gas turbine and the ﬂows of gas and steam have a decreasing tendency when the ambient temperatures are hotter.

When the gas turbine load decreases, the exhaust temperature of the gas turbine goes down. As the mass ﬂow of gas contains less energy the steaming potential in the HRSG decreases.

16 2 Combined Cycle Power Plants

Figure 2.14: Energy transfer diagram in an HRSG of a combined cycle power plant, [4]

2.4 Combined Cycle Power Plants A Combined Cycle Power Plant, CCPP, is the combination of two cycles: Brayton and Rankine. The Brayton cycle works as the topping cycle and the Rankine cycle as the bot- toming, being the energy transferred from the gas cycle to the steam cycle. The HRSG is the component which transfers the energy from one to the other cycle. Thermal eﬃciencies of the combined cycles can reach as high as 60%, Boyce [3]. In the typical combination the gas turbine produces about 65% of the power and the steam turbine about 35%, while unit thermal eﬃciencies of the gas turbine and the steam turbine are between 30%-40%.

The condensate entering the HRSG goes through a deaerator where the gases from the

17 2 Combined Cycle Power Plants

Figure 2.15: HRSG performance versus ambient temperature. Gas ﬂow shown has a multi- plication factor of 0.1, [5] water or steam are removed. This is important because a high oxygen content can cause corrosion in the pipes and the rest of components in contact with the water/steam. The deaerator is normally placed on top of the feedwater tank and the process occurs when the water is sprayed and then heated, thus releasing the gases.

Deaeration also takes place in the condenser and sometimes this could lead to not utiliz- ing a separate deaerator/feedwater tank, and the condensate being fed directly into the HRSG from the condenser.

The economizer is used to heat the water under its saturation temperature. The risk of generating steam in the economizer has to be taken into account because that can block the ﬂow. To prevent the appearance of steam in the economizer a feedwater control valve can be installed for keeping the pressure high and avoiding the steaming.

The steam turbines in most of the large power plants are at minimum divided in two sections: the High Pressure Section (HP) and the Low Pressure Section (LP). In some plants, the HP section is as well divided into a High Pressure Section and an Intermediate Pressure Section (IP). The HRSG is divided in the same sections as the steam turbine. The LP steam turbine’s performance is further dictated by the condenser backpressure, which is a function of the cooling and the fouling.

The efficiency of the steam section in many of these plants varies from 30%-40%. To ensure that the steam turbine is operating in an efficient mode, the gas turbine exhaust temperature is maintained over a wide range of operating conditions. This enables the HRSG to maintain a high degree of effectiveness over this wide range of operation.

In a combined cycle plant, high steam pressures do not necessarily imply a high thermal eﬃciency. Expanding the steam at higher pressure causes an increase in the moisture con-

18 2 Combined Cycle Power Plants

Figure 2.16: Energy Distribution in a Combined Cycle Power Plant, [3] tent at the exit of the steam turbine. The increase in moisture content creates major erosion and corrosion problems in the later stages of the turbine. A limit is set at about 10% moisture content, Boyce [4].

One advantage of working with high steam pressures are that the volume ﬂow of steam is reduced. That means that the sections of turbines, condenser, pipes and the rest of components can be smaller and the cost decreases.

However, increasing the steam temperature at a given pressure lower the steam output of the steam turbine slightly. This occurs because of two contradictory effects: first the increase in enthalpy drop, which increases the output; and second the decrease in flow, which causes a loss in steam turbine output. The second effect is more predominant, which imply the lower amount of steam. Besides, lowering the temperature of the steam also increases the moisture content.

Increasing the pressure of any section (LP, IP and HP turbines) will increase the work output of the section for the same mass ﬂow. However, at higher pressure, the mass ﬂow of the steam generated is reduced.

19 3 Concentrating Solar Power Plants

3 Concentrating Solar Power Plants

Solar energy is the energy generated by the sun which is converted in useful energy by the human being for heating or electricity generation. Every year the sun supplies 4 times more energy than we consume, so its potential is almost unlimited.

The technology used in thermal power plants is the same used in conventional power plants, except that the heating source is the Sun. The heat energy is used to drive a steam turbine and to produce electricity with generators coupled to the turbines. In the conventional power generation, the heat energy comes from combustion of fossil fuels or from nuclear ﬁssion. Un- like traditional power plants, concentrating solar power systems produce a source of energy without emissions and the only impact is the land use. Other beneﬁts of concentrating solar power plants include low operating costs and an increase in energy independence from foreign oil imports.

Solar energy, in contrast with fossil fuels, is not available around the clock and the intensity of the available energy in a deﬁned point of the Earth depends on the day of the year, the hour and the latitude. Besides, the quantity of energy that can be obtained depends on the orientation of the receiver mechanism. The gaps can be ﬁlled in two ways: switching to fossil fuel combustion or storing the colected heat energy for when the Sun is not available.

Concentrating solar power plants produce electric power by converting the energy of the sun into high-temperature heat using various mirror conﬁgurations which concentrate the rays of the Sun for obtaining a high temperature. At least 300◦C are necessary in order to generate power eﬀectively and economically. For reaching this is essential to have a high percentage of direct solar radiation. This is more the case in the Sun Belt between the 35th northern and 35th southern latitudes.

Concentrating solar power (CSP) plants consist of two parts: one that collects solar energy and converts it to heat (solar ﬁeld), and another that converts heat energy to electricity (power island). The power gained from sunlight can be increased if the light is gathered and concentrated on a single point and the heat is then channeled through a conventional generator. For concentrating the Sun’s rays there are two possibilities: concentrating the radiation in a ﬁxed point (here the concentrators have to follow the Sun by moving along two axes) or using linear concentrators and only need to move along one axis in order to follow the Sun.

3.1 Solar Thermal Concentrating Collectors In practice the most used concentrators are linear concentrators because they are more suitable for assembly on a large scale and less costly to construct. The diﬀerent types of linear concentrators are explained in the following.

3.1.1 Parabolic-Trough Power Plants The ﬁrst commercial CSP were parabolic trough systems installed in the United States in the 1980’s and are the most proven, developed and commercially-ready CSP technologies.

20 3 Concentrating Solar Power Plants

Parabolic trough technology is a clean and mature solar power solution with years of success- ful power generation behind it. The technology has been improving steadily for the last 30 years, and modern troughs operate more eﬃciently and at lower cost. Today, there is more than 700 MW of CSP trough power in operation around the world, with 400 MW under construction and around 20 GW in development.

The parabolic trough collector is a linear concentrator which is moved along one axis and focuses the Sun’s energy to a fluid carrying receiver tube situated at the focal point where the mirror concentrates all the sunlight. The tube is filled with a thermo-oil as heat transfer medium and it is heated when it passes through the solar collector and then flows into a heat exchanger where the steam is generated. The latest parabolic trough collectors use synthetic oils and they can reach the temperature of 350◦C. Due to the limited temperature in thermo-oil systems and also the high cost of the thermo-oil, next generation of CSP aims at using direct steam generation. We will use water in our study of a CCPP with Solar Boosting.

Figure 3.1: Operating Scheme for Parabolic Trough Technology, [9]

The troughs are placed in parallel rows to form a collector ﬁeld. The rows lined-up along a north-south axis can track the sun’s trajectory during the day, ensuring that the sunlight is always focused on the receiver tube. Some trough systems incorporate thermal storage to produce electricity even when the sun is not shinning.

The actual dimensions of the parabolic trough collectors are: the opening width have 5.71 m and the collector length have 99 m. The absorber pipe is made of steel, covered with an optically-selective surface coating and inserted in an evacuated glass tube. It absorbs the radiation from the solar spectrum eﬀectively keeping the thermal losses to the ambient at minimum by re-radiating only a small amount of radiation energy. Thick glases with a

21 3 Concentrating Solar Power Plants low content of iron are used to built the mirrors. German organizations have developed the Eurotrough collector, which is lighter and stiﬀer, and is less cost to produce, to assemble and to maintain.

Figure 3.2: Parabolic trough solar collectors at Kramer Junction in the Mojave desert in California, [19].

In Europe, two power plants called Andasol I and II and situated in Granada, Spain, are currently in operation using the parabolic trough technology and delivering to the electric grid 50 MWe output power each. In California, the nine SEGS (Solar Electricity Generating System) have three generations of parabolic trough collectors in service, Wengenmayr [8].

Linear Fresnel Collectors are a similar design which uses a long, narrow, shallow-curvature or ﬂat mirrors to focus the light onto a linear absorber positioned above the mirrors.

3.1.2 Dish Stirling Systems The dish concentrator is a parabolic silvered dish, which by moving along two axes focuses the radiation of the Sun onto a single point. At the focus point the receiver is situated and there is as well a thermal engine directly connected to it. These systems work as an independent power generation, being unconnected to a power grid.

Their principal advantage is a very high eﬃciency over 30%, which is a result of the combination of a nearly paraboloid concentrator with an excellent thermal engine. When the Sun is not shining the dish stirling can be fed with fuel is satisfy the power demand. The maximum size of the dish stirling is limited by the wind forces, which deform the reﬂecting surface. The maximum surface is 100 m2 and the maximum electric power output is 10kW.

22 3 Concentrating Solar Power Plants

Figure 3.3: Solar Power Group’s Fresnel mirror test rig at the Plataforma Solar de Almeria in Spain,[19].

The costs of generating electricity are still above those of parabolic trough power plants. However, cost may decrease in the near future.

Figure 3.4: Dish Stirling Operating Scheme, [9]

3.1.3 Central Receiver System The central receiver system divides the big surface of the dish stirling in smaller concentrator mirrors called heliostats. The heliostats focus the rays of the Sun onto a common point sit-

23 3 Concentrating Solar Power Plants uated on a central tower, where the receiver collects the heat. The heliostats are ﬂat mirrors and they can not concentrate as much as the ideal paraboloid does.

Large central receiver systems with thousands of heliostats, each with 100 m2 of reﬂect- ing surface, require towers of 100-200 m high and they can collect several hundred MW of solar radiation power.

The ﬁrst plant with central receiver system called PS10 is situated in Sol´ucar Platform in Sevilla, Spain. Its started operation in 2007 with 11 MW of output power. The second and biggest in the world, PS20, is since 2009 in operation with 20 MW of power output.

Figure 3.5: Aerial Photography of Sol´ucar Platform with the PS10 and PS20 Receiver Tower Systems

Figure 3.6: Tower Receiver Operating Scheme, [9]

In the ﬁrst test of this technology the doubt was the heat transfer medium to use. It ﬁrst appeared attractive to use the superheated steam itself which avoids the intermediate heat exchangers and enables a direct connection to the steam turbines. However, the problem

24 3 Concentrating Solar Power Plants was how to control the generation of superheated steam with changing radiation conditions. Besides, storing the heat energy was almost impossible without having a big amount of heat losses. Therefore, the PS10 Solar Power Plant is based on saturated steam at moderate temperature and pressures to avoid these problems.

Other possibility is the use of alkali-metal salts as heat medium transfer. They have two advantages: the good heat transfer properties and the possibility of storage at low pressures in tanks. However, the high melting point makes necessary electrical heating the pipes to avoid freezing out of the salts with the pipe blockage as result.

3.2 Parabolic Trough Power Plant Conﬁgurations 3.2.1 Solar Mode In these systems the only heat source which moves the steam turbine is the Sun. As the Sun does not radiate continuously the heat source is not stable and systems for thermal storage are necessary. In summer, the average operating hours are 10-12 hours, which means the remaining time the plant can only be operated using the stored energy.

Since the Sun rises in the morning the solar ﬁeld starts delivering heat to the thermal cycle. During the radiation peak, the excess heat is stored in the thermal storage system which will be in charge of delivering the necessary heat during the period when the Sun’s radiation is not enough.

The diﬀerences between the available thermal storage systems are basically the storage medium used. In the following we will describe the most common storage mediums: • Salt: Sodium-nitrate salts and Potasium-nitrate salts are cheap materials for storage systems. They have a high transmission coeﬃcient and can be stored in big salt tanks. The problem is their high melting point which needs electrical heating of the piping for avoiding the blockage for example during system start up.

3.2.2 Direct Steam Generation In a parabolic trough collector, the oil as a heat transfer ﬂuid is heated by concentrated solar radiation. The thermal energy contained in the oil is transferred to the Rankine cycle where the power is generated. One of the limitations is the chemical stability of the synthetic oil when it is heated at high temperatures, being the limit about 395◦C. That limits the maximum temperature of the steam in the Rankine cycle being not possible to reach higher eﬃciencies.

One of the possible alternatives is Direct Steam Generation (DSG)in the collector field. This configuration consists of evaporating and superheating the water directly in the solar field which makes the oil-water heat exchangers unnecessary. With the DSG the steam can reach temperatures from 400◦C to 550◦C increasing the thermal efficiency of the Rankine cycle. The optimization and demonstration of these technology components has been done in the a 700 m collector loop in the Plataforma Solar de Almer´ıa, [10].

Three diﬀerent regimes of operation are possible:

25 3 Concentrating Solar Power Plants

• Once Trough System: preheates, evaporates and superheates the feed water. This system is the simplest and cheapest but the control of the temperature in the receiver tube is complex due to the inhomogeneous distribution of temperature on the tube circumference.

• Injection System: the water is injected in several points of the receiver tube. This system presents the problem of a complex measurement and control operation.

• Recirculation System: in this system the collector tube is divided in two sections. In the ﬁrst section the water is preheated and evaporated, while in the second the water is superheated. In between the two sections there is a water-steam separator, where the water content in the mixture is separated and sent back to the solar collector inlet.

3.2.3 Integrated Solar Combined Cycle (ISCC) The Integrated Solar Combined Cycle (ISCC) consists of a conventional Combined Cycle Power Plant (CCPP) which is combined with a parabolic trough solar field. The solar field produces superheated steam which is fed into the heat recovery steam generator (HRSG) of the combined cycle allowing the increase of the themal efficiency of the steam cycle in the CCPP.

The beneﬁts of employing this technology are to overcome some problems related with

Figure 3.7: ISCC Operating Scheme the start up and shut down in solar power plants, reduce the capital costs and improve the solar-to-electricity eﬃciency.

26 4 Modeling and Simulation with EBSILONr Professional

4 Modeling and Simulation with EBSILONr Professional

In this chapter several modeling processes of a Combined Cycle Power Plant will be described. The software used for modeling is EBSILONr Professional 8.0, which we will call Ebsilon. Ebsilon is a mass and energy balance calculation program for thermodynamical cycles. With Ebsilon we will be able to simulate the performance of a combined cycle power plant in design and partload conditions, which is adequate for analyzing its performance under several loading conditions.

In the following, we will shortly describe the features and tools available in Ebsilon.

4.1 Basics of the Ebsilon Software EBSILON is the abbreviation for ”Energy balance and simulation of the load response of power generating or process controlling network structures” and is suitable for nearly all stationary thermodynamic model request coming out of energy cycles or plant schemes.

Ebsilon permits the balancing of components, individually or in groups, as well as sub- systems integrated in bigger systems, without taking into account whether these components or systems form a closed or an open cycle. The model structure of Ebsilon is based on:

• standard components, which are used for modeling common power plants,

• programmable components for modeling complex power plants processes with user- deﬁned behaviour.

The data basis of Ebsilon is made up of:

• IAPWS-IF97 or the IFC67 steam table,

• cp-polinomial for air/fuel gas.

Ebsilon is a variable program system, by means of which you can balance all occurring power plant circuits using a closed solution based on a sequential solution method.

The cycles are constructed from objects. The object types can be:

• Components

• Pipes

• Macros (such as the gas turbines from the library)

• Value Crosses

• Text Fields

• Graphical elements

27 4 Modeling and Simulation with EBSILONr Professional

• OLE Objects

Components and pipes form cycles, while value crosses, text ﬁelds, graphical elements and OLE objects are used for displaying the results, such as comments and explanations.

In a short description, the basic control elements and tool bars are:

• Standard toolbar,

• Component bar, for selecting a component from a category,

• Component wizard bar, for accessing components classiﬁed by numbers,

• Ebsilon bar, for starting simulations,

• Zoom bar, for zooming in the model and ﬁnding objects.

4.1.1 Data introduction After deﬁning the topology of the cycle, it is time to deﬁne the values that characterize it. There are two ways of doing that:

1. Some components can deﬁne or calculate values in aﬃliated pipes, like e.g. the steam generator, as shown in Figure 4.1.

2. The other possibility is to set the value directly on the pipe using the ”Measured/Gen- eral value input”, as shown in Figure 4.2.

28 4 Modeling and Simulation with EBSILONr Professional

Figure 4.1: Deﬁning data on component’s pipes

Figure 4.2: Deﬁning data on the pipe

4.1.2 Calculation Modes After building the cycle we set the design reference by defining the nominal conditions. Once the calculation in nominal conditions is done, the calculated parameters are memorized and we can start analyzing the cycle in different conditions. In the following we present a shortly description of the different available modes for working in Ebsilon. • Design Model: in this model the power plant is built taking into account the design conditions. After defining the features of the plant, it is simulated and the parameters are calculated automatically. Some fixed parameters, as the surfaces of the heat exchangers, will be memorized. • Partload Model: Ebsilon provides the possibility to create different profiles from one model, each one has the same topology but its own parameters and results. The partload operation is used when we want to simulate the performance of the plant working

29 5 Model of the Gas Turbine GE 9FA

under different load situations. The new profile Partload has the same parameters as the root profile (Design) but calculates in off-Design mode. All the input values are inherited from the parent profile. • Validation Model: that model is used for controlling the performance of existing plants. Weak points can be detected and from that derives that quality statements about power plants can be derived. • Optimization Model: the optimization model is used for the variation of power plant parameter and what-if calculations. Different strategies of optimization can be applied as well as a combination with the validation model. In our analysis of Solar Retrofit in Combined Cycle Power Plants. We will use the Design and Partload models in studying the impact caused for feeding the plant with energy coming from a solar field.

5 Model of the Gas Turbine GE 9FA

5.1 Gas Turbine GE 9FA in Design Conditions The chosen turbine for the design of our combined cycle power plant is the General Electric’s GE 9FA. The nominal conditions for the gas turbine are presented in the next Figure 5.1:

Figure 5.1: Nominal conditions of the GE 9FA in a simple cycle or in a CCPP, [13]

In the construction of the cycle of the GE 9FA Turbine, the selected components are a compressor, a combustion chamber and a turbine. The conditions established during the design

30 5 Model of the Gas Turbine GE 9FA process are the ISO conditions: 1.013 bar and 15◦C for ambient air.

In nominal conditions, the turbine has a pressure ratio of 17 and is crossed by a mass flow of 641 kg/s. The temperature of the gas flow after the expansion is 602◦C at the outlet. Using a ”general input value” we set the nominal exhaust temperature and the mass flow, while the pressure is sett in the turbine component. The general input value has to be situated after the turbine, because we only know the exhaust temperature. After writing all data and simulating the performance of the turbine, we obtain that 627.04 kg/s of air are necessary for reaching the nominal conditions in the gas turbine.

GE Turbine 9FA 50Hz 17.002 32.681 ISO Conditions 15.000 13.926

17.002 436.334 17.000 1514.160 P bar H kJ/kg 1278.077 641.000 T °C M kg/s 420.710 627.074 260017.690 kW

270851.761 kW

1.013 15.155 1.013 666.952 15.000 627.074 602.000 641.000

Figure 5.2: Model of GE 9FA in Design Conditions

Once we know that information we are able to set the air conditions at the entrance of the gas turbine. Using a ”boundary value input” we set the conditions of the air (1.013 bar, 15◦C and 627.04 kg/s) and simulate the performance again. The results are the same that in the step before, we are reaching again the nominal conditions for the gas turbine. The reason of doing this last step is because we will control the turbine performance in partload by means of the amount of air which enters the compressor and the mass of fuel burnt in the combustion chamber. Now, it is possible to set the new parameters of the air in partload by introducing the values in the ”boundary value input” at the entrance of the gas turbine.

Comparing the data of the Figure 5.2 and Figure 5.1 we can conclude that the performance of model of the GE 9FA Turbine represents approximately the performance of the turbine in reality. Both the exhaust temperature and the gas flow coincide with the data given for General Electric and the only difference is the power output, which is higher in our model. That could be a consequence of a different air ratio which could result as a different turbine inlet temperature, different isentropic efficiencies in turbine and compressor or the electrical efficiency of the generator.

The previous results show the gas turbine behaviour in nominal conditions. In the following

31 5 Model of the Gas Turbine GE 9FA we will discuss the behaviour of the same turbine when the conditions of the surroundings change and when the net needs less power than the delivered power in nominal conditions.

5.2 Gas Turbine GE 9FA in Off-design The gas turbine working in nominal conditions means that the turbine works in the same conditions for which it was designed. Partload conditions are the conditions in which the turbine works under different specifications. Partload can be reached when the ambient conditions changes, the temperature, or when the turbine is delivering less power than the nominal power. The objective of this study is to establish the performance of the gas turbine in the first case, when the air temperature differs from the ISO temperature, although we will also briefly describe the performance when less than nominal power is enough.

5.2.1 Variation of Ambient Temperature The variation of ambient temperature produces a change in the turbine behaviour because the air entering the turbine changes its properties, such as temperature and humidity.

For our study we will suppose a variation of 5% in humidity (φ) when the temperature changes 15◦C. Besides, the density (ρ) changes when the temperature does in a way that, the density is inversely proportional to the temperature.

This has to be taken into account because in the different seasons of the year, the ambient temperature raises, being higher than 15◦C, or decreases and the air density is different as well. That means that the same turbine working at a fixed place will deliver more or less power depending on the season of the year.

In Europe, in summer normally the ambient temperature is higher than 15◦C and that means the air density is lower. As we know, the air density is inversely proportional to its volume, so the volume occupied for the same amount of air will be bigger. Thus, the amount of air entering the compressor will be lower, due to the fact that compressor has a fixed volume capacity, and the gas flow passing through the turbine is lower too. That means that the same turbine working in summer has lower power output, because the expanded mass flow is smaller. If we think that in summer the demand of electricity is not lower than in winter, although no energy for heating is required, the same turbine working in summer conditions will not be enough for satisfying the energy demanded. For this reason, it is necessary to analyze how much can vary the turbine efficiency depending on the ambient temperature.

In winter, for colder temperatures than 15◦C, the opposite occurs. The density will decrease and the compressed mass of air will be bigger. That changes the performance of the turbine in a way that in winter delivers more power, if there is no limit for the turbine inlet temperature.

In the following we will analyze three representative cases for the partload behaviour, 30◦C, 45◦C and 0◦C.

First of all, we have to calculate the variation of density and air ﬂow for each case. Knowing

32 5 Model of the Gas Turbine GE 9FA the decrease percentage in the humidity when the temperature rises, we can calculate the resulting density at the required temperature which let us calculate the mass ﬂow of air entering the compressor of the gas turbine.

As example, the calculations for the case at 30◦C are described in the following:

TN 3 ρ30 = ρN = 1.147kg/m (5.1) T30

ρ30 m˙30 =m ˙N = 596.03kg/s (5.2) ρN

The density values obtained in each case are showed in the Table 5.2.1:

◦ 3 Tamb ( C) ρ (kg/m ) φ (%) 0 1.273 105 15 1.207 100 30 1.147 95 45 1.11 90

At 30◦C, we get a mass of air of 596.03 kg/s, which is smaller than in nominal conditions as we said. The new air ﬂow has to be set at the compressor entrance using the ”boundary value input” as we explained before. The temperature has to be changed too.

In every case of partload we will have now new conditions of temperature and mass ﬂow. Ebsilon calculates by itself the amount of fuel which has to be burnt in the combustion chamber. The parameter which determines the amount of fuel is the air ratio.

The air ratio (called ALAM in Ebsilon and represented in tables as λ) is a parameter in the combustion chamber which enables the regulation of the maximum temperature of the cycle. The air ratio is deﬁned as the ratio between the mass of air and the stoichiometric mass of air for a known fuel ﬂow.

By changing the air ratio, we give the order to the combustion chamber to accept more or less fuel. That means that the maximum temperature of the cycle, after the combustion chamber, is changing. This is interesting because in our case that we are studying we want either higher or lower temperatures. That can be regulated by varying the air ratio. If we increase the air ratio we accept more air in reference to the stoichiometric air and the maximum temperature of the cycle decreases. The exhaust temperature of the turbine decreases as well. If the air ratio decreases, the temperatures increase.

For determining the exhaust temperature of the turbine, we have followed the next criteria described:

1. When the ambient temperature is hotter than in nominal conditions (30◦C and 45◦C), the temperature in the exhaust of the turbine remains constant.

33 5 Model of the Gas Turbine GE 9FA

2. At colder temperatures than in nominal conditions (0◦C) that the maximum temperature in the cycle remains constant. The temperature before entering the turbine does not vary.

Figure 5.3 shows the evolution of the exhaust and inlet temperature of the gas turbine according to the criteria followed for establishing the temperatures in partload.

Tin (ºC) Texh (ºC)

TIT (ºC) TOT (ºC) 1290 605

600 1280 595 1270 590 1260 585

1250 580

575 1240 570 1230 565 1220 560

1210 555 0 15 30 45 Atmospheric Temperature

Figure 5.3: Temperature Evolution in Oﬀ-design

We already know that in a combined cycle power plant, the energy in the exhaust gases of the gas turbine is transferred to the steam cycle by using a HRSG. The exhaust gases of the gas turbine have a big amount of energy when they are leaving the turbine with a temperature over 600◦C. The parameters which deﬁne the HRSG are set for the design case, that means that the surfaces of the heat exchangers are designed according to these conditions. At other conditions they are of course constant. For adequate performance of the combined cycle power plant the conditions of the steam cycle should remain constant. In consequence the outlet temperature (TOT) of the gas turbine should be as constant as possible independent from the ambient temperature.

In the Table 2 and the Table 3 the diﬀerent values of mass ﬂows, temperatures and air ratios obtained in each case are shown:

34 5 Model of the Gas Turbine GE 9FA

Figure 5.4: Variation of gas turbine performance in T-s diagram depending on the ambient temperature, [7]

◦ Tamb ( C) m˙ air (kg/s) m˙ fuel (kg/s) m˙ gas (kg/s) 0 661.5 15.9 677.0 ISO 627.1 13.9 641.0 30 596.0 12.4 608.5 45 576.0 11.4 588.4

Table 2: Mass ﬂows in Oﬀ-design

As we can see, the power of the gas turbine is higher when the ambient conditions are colder due to the fact that a bigger amount of gas is expanded in the turbine and the pressure ratio is higher than in nominal conditions. That makes the gas turbine power output higher in winter, while in summer it decreases.

5.2.2 Variation of fuel and air ﬂows When an amount of power smaller than in nominal conditions is enough for satisfying the demand, the gas turbine can works in partload as well.

In ﬁrst place, the pursued objective is to determine how much power we need. As we said, we want to obtain the variation in power but always taking into account that after the gas turbine there is a heat recovery steam generator and the exhaust temperature in the gas turbine has to remain constant. For adjusting the exhaust temperature we will vary again the air ratio.

In this case the criteria followed establishes a constant exhaust temperature while the range of power is between the 50% of the load and nominal conditions. For obtaining a constant exhaust temperature, the air ratio has to be changed in the same way that we have already explained when the gas turbine works in an ambient with changing temperatures. For obtaining the required power, the mass of air has to be changed as well, taking into account

35 5 Model of the Gas Turbine GE 9FA

◦ ◦ ◦ Tamb ( C) pin (bar) λ Power (MWe) TIT ( C) TOT ( C) 0 18.0 2.462 318.63 1278.0 573.0 ISO 17.0 2.595 260.02 1278.1 602.0 30 16.0 2.762 213.53 1253.4 602.0 45 15.4 2.916 183.85 1238.6 602.0

Table 3: Parameters in Oﬀ-design that the smaller amount of air the less power is delivered. Depending on the required power, Ebsilon calculates by itself the amount of fuel necessary for doing that the gas turbine delivers exactly that amount.

The Figure 5.5 shows the evolution of the gas ﬂow and exhaust temperature when the performance of the gas turbine varies from 30% to 110% of the nominal power.

• The ratio Mgas/Mgas0 represents the decrease of the gas ﬂow with regard to the nominal gas ﬂow.

• The ratio Texh/Texh0 represents the variation of temperature with regard to the exhaust temperature in design conditions.

Mgas/Mgas0 Texh/Texh0

110%

105%

100%

95%

90%

85%

80%

75%

70%

65%

60% 30% 40% 50% 60% 70% 80% 90% 100% 110%

Paprox /P 0

Figure 5.5: Temperature and Mass flow Behaviours in Off-desing for a specific power

As we can see in the Figure 5.5 less mass goes through the turbine when we move into partload operation. That means that the air ﬂow and the fuel ﬂow decrease.

36 5 Model of the Gas Turbine GE 9FA

Figure 5.6: Comparison between the T,s diagrams for the gas turbine accepting less air

In the Figure 5.6 the cycle 1-2’-3’-4’ represents the gas turbine working with less air. In the graphic, we can see that when the air ﬂow is small the compression process is shorter and the combustion is longer until reaching the maximum temperature in the cycle. This new ”longer” combustion in process 2’-3’ means that the amount of fuel burnt is bigger than in process 2-3. In a real gas turbine the mass of air entering the compressor is controlled by changing the orientation of the blades at the entrance (inlet guides vanes).

Although in Table 4 the amount of fuel is decreasing at the same time as the amount of air, the fuel decreases slower compared with the variation of air ﬂow. The decrease of the amount of air entering the compressor leads to a constant outlet turbine temperature, as we can see in Figure 5.5.

In Table 4 the diﬀerent values obtained for every parameter are shown in each case of oﬀ-design:

• The parameter Partload represents the ratio between the necessary power and the nominal power.

• The parameter Paprox is the power that in reality is obtained in the turbine in each case.

• The ratio Paprox/P0 represents how close we are of obtaining the partload percentage necessary.

As an example we will explain the case of 80%:

The amount of power obtained in nominal conditions is 270852 kW. In this case we need the

37 5 Model of the Gas Turbine GE 9FA

80% of the nominal power, which is 0, 8·270852 = 216682, 21 kW. For reaching that amount, we will vary the air ratio and the mass of air entering the compresor. That also produces that the mass of fuel changes. At the end we obtain a value similar to the power needed, 216683 kW. This value is called Paprox and the ratio Paprox/P0 represents the exactitud with which the delivered power is obtained. If the Paprox/P0 coincides with the Partload needed, then the deliverd power is exactly the required amount. In that case of 80% the ratio Paprox/P0 is exactly 80%, which means that the numbers obtained are correct.

In Figure 5.7 the model of the gas turbine working at the conditions of the case described is showed.

GE Turbine 9FA 50Hz 15.099 32.681 ISO Conditions 15.000 11.988

80% of Nominal Power 15.099 417.655 15.097 1445.996 PGE Turbine bar H 9FA kJ/kg 50Hz 1228.503 578.568 T °C M kg/s 403.385 566.580 15.099 32.681 ISO Conditions 15.000 11.988 200874.938 kW

80% of Nominal Power 15.099 417.655 15.097 1445.996 P bar H kJ/kg 1228.503 578.568 T °C M kg/s 403.385 566.580 G 200874.938 kW 216682.610 kW

1.013 15.155 1.013 665.534 15.000 566.580 216682.610 kW 602.007 578.568

1.013 15.155 1.013 665.534 15.000 566.580 602.007 578.568

Figure 5.7: Gas Turbine working at 80% of nominal power

Although in Table 4 the amount of fuel is decreasing at the same time as the amount of air, the fuel decreases slower compared to the variation of air ﬂow.

5.2.3 Variation of Efficiency in Components As we explained before, the gas turbine efficiency depends only on the pressure ratio and the nature of the working fluid. Working in partload conditions means changes in the pressure ratio of the gas turbine and in its power output. But also the other components of the gas turbine working in part load suffer a change in their performance which has to be taken into account in the overall efficiency of the gas turbine.

38 5 Model of the Gas Turbine GE 9FA

◦ Partload P0 (kW) Paprox (kW) m˙ air (kg/s) m˙ fuel (kg/s) m˙ gas (kg/s) λ Texh ( C) 110.00% 297937 297932 625.0 15.1 640.1 2.380 643.1 100.00% 270852 270852 627.1 13.9 641.0 2.595 602.0 90.00% 243767 243768 597.9 13.0 610.9 2.657 602.0 80.00% 216682 216683 566.6 12.0 578.6 2.724 602.0 70.00% 189597 189597 532.6 11.0 543.6 2.798 602.0 60.00% 162512 162512 495.3 9.9 505.2 2.881 602.0 50.00% 135426 135426 453.9 8.8 462.7 2.976 602.0 40.00% 108341 108342 447.0 7.4 454.4 3.459 540.1 30.00% 81256 81304 444.5 6.1 450.6 4.220 466.6

Table 4: Variation of Parameters in Oﬀ-design when the power is speciﬁed

Paprox/P0 Mgas/Mgas0 Texh/Texh0 110.0% 99.7% 106.8% 100.0% 100.0% 100.0% 90.0% 95.4% 100.0% 80.0% 90.4% 100.0% 70.0% 84.9% 100.0% 60.0% 79.0% 100.0% 50.0% 72.4% 100.0% 40.0% 71.3% 89.7% 30.0% 70.9% 77.5%

Table 5: Relative Parameters in Oﬀ-design

In Ebsilon compressors and turbines have an established default value of isentropic efficiency. The isentropic efficiency in a compressor or a turbine is a comparison between the real power obtained or consumed and the isentropic case. The default isentropic efficiency for turbines is 0.9 and for compressors is 0.85 and in partload that value is defined by some correction curves. The variation of isentropic efficiency is directly proportional to the change of mass flow which is going through the compressor or turbine.

The Tables 6 and 7 show the respective correction curves:

m/˙ m˙ N ET AI/ET AIN 0 0.85 0.4 0.9 0.7 0.95 1 1 1.2 1.1

Table 6: Variation of Eﬃciency in the Turbine with the mass ﬂow

The value of the isentropic eﬃciency is represented in Tables 6 and 7 by the name of ETAI.

39 5 Model of the Gas Turbine GE 9FA

m˙ 1/m˙ 1N ET AI/ET AIN 0 0 0.4 0,9 1 1 1.2 1.1

Table 7: Variation of Eﬃciency in the Compresor with the mass ﬂow

5.3 Oﬀ-load Model of HRSG, Steam Turbine and Condenser In this chapter we pretend to analyze how the parameters in every component variate when the operation of the steam cycle is not in design conditions.

Heat Exchangers

The heat transfer capacity of a heat exchanger variates with the change in mass flow trough the heat exchanger. The transfer surfaces in the heat exchanger are defined in design conditions and remain constant in every mode of operation. The variation of the product of the heat transfer capacity and the transfer surface is defined by some default correction curves implemented in Ebsilon.

A heat exchanger is a component where two fluids transfer heat between them. The Figure 5.8 shows a section of a heat exchanger in Ebsilon software. Independent of water or steam (1-2) the variation of the heat transfer capacity of the heat exchanger for each fluid is represented in Table 8. In Table 9 the gas side (3-4) is represented. As we can see, it is assumed that the heat transfer capacity remains constant in the gas pipe of the heat exchanger when the mode of operation is off-load.

Figure 5.8: Section of a heat exchanger in Ebsilon

Condenser

The heat transferred in the condenser also variates with the mode of operation. Although the condenser is also a heat exchanger we only study the pipe in which the steam is condensed to water, the other pipe is not interesting because it is suppose that the condenser can use as much as it needs for condesing the steam. The Table 10 shows the variation of heat transfer capacity with the mass of steam condensed in oﬀ-load.

Steam Turbine

In steam turbine, the pressure variates in accordance with the Stodola’s law which deﬁnes a variation in pressure proportional to the variation of mass ﬂow expanded through the steam

40 5 Model of the Gas Turbine GE 9FA

M1/M1N kA1/kAN 0 0.5 0.4 0.6 0.5 0.7 0.6 0.8 0.7 0.85 0.8 0.9 0.9 0.95 1 1 1.1 1.04 1.2 1.07

Table 8: Variation of Heat Transfer Capacity with the mass ﬂow on the steam/water pipe.

M3/M3N kA2/kAN 0 1 0.4 1 0.5 1 0.6 1 0.7 1 0.8 1 0.9 1 1 1 1.1 1 10 1

Table 9: Variation of Heat Transfer Capacity with the mass ﬂow in the gas side

M1/M1N kA1/kAN 0 0.5 0.4 0.6 0.5 0.7 0.6 0.8 0.7 0.85 0.8 0.9 0.9 0.95 1 1 1.1 1.04 1.2 1.07

Table 10: Variation of heat transfer capacity with the condensed mass flow. turbine . The correction curve defines in this case a correction factor, with which the pressure calculated according to the Stodola’s expansion law is multiplied. This simplifies the pressure adaptation to the real plant states. The correction curve is shown in Table 11.

41 6 Modeling a Single Pressure CCPP

M1/M1N P 1/P 1St 0 1 0.2 1 0.4 1 0.5 1 0.6 1 0.7 1 0.8 1 0.9 1 1 1 1.1 1 1.2 1

Table 11: Stodola Correction for the Steam Turbine

6 Modeling a Single Pressure CCPP

The simplest steam cycle consists of a steam turbine, a condenser, the heat exchangers of the one level pressure heat recovery steam generator and the pumps. This is the single pressure cycle (1P) and the heat recovery steam generator consists of two economizers, the evaporator, two superheaters and one reheater.

The performances of the steam turbine and the condenser have been explained in the chapters ”Steam Cycle” and ”Types of Condensers”, but it is important to notice that in our CCPP the steam turbine will have always three stages of expansion although the HRSG has only a single pressure level. The reason of three stages in the steam turbine is because we have reheat and a deaerator and we need some opening point for taking the steam. In the reheat the steam goes again through a heat exchanger where it is heated and recovers its energy and the deaerator needs to extract some steam from the steam turbine at low pressure.

By superheating the steam leaving the evaporator we ensure that the saturated steam converts into dry steam and there is no water droplets in the steam ﬂow. The superheater also rises the steam temperature at the inlet of the steam turbine which allow us to obtain more power output.

The reason of having two superheaters in the HRGS is because we need to control the temperature. Feed water is injected for controlling the superheat temperature. Although the pipe connection is made in the cycle, we will not use the feed water injection for controlling the superheater temperature in the modeling of the one pressure CCPP. This temperature will be controlled in our model by using the settings in the components.

The reheat process has been explained in the chapter ”Improvements for increasing the work output” of the gas turbine. The pipe which enters the reheater is called the cold reheat pipe. In this pipe there is another connection for injecting feed water in case we want to control the reheat temperature. As in the case of superheaters we will not use it, being the mass ﬂow through the pipe 0 kg/s. The reheat temperature will be controlled by using the setting in the reheater pipes.

42 6 Modeling a Single Pressure CCPP

The performance of the evaporator has also been explained in the chapter ”Types of HRSGs”. As a shortly description we should say that the single pressure cycle has only one evaporator which deﬁnes the one level pressure HRSG. In the evaporator the water is converted into saturated steam at constant pressure and temperature. The evaporator has a drum where the water is separated from the steam and recirculated for being evaporated, while the steam is leaving the evaporator for entering the superheaters.

The economizer consists of a heat exchanger where the water is heated before entering the evaporator. The purpose of this component is to save energy by preheating the water before the evaporator. In the single pressure cycle we have two economizers. That is because we can extract water between the economizers and inject it as in the cold reheat pipe.

The steam cycle also has a deaerator. The deaerator is the component which guarantees that there is no air and other dissolved gases going from the feed water to the HRSG. Dissolved oxygen in boilers can cause corrosion damage in the steam systems and form- ing oxides in metal piping. The water which enters the deaerator comes from a preheater where its temperature is raised being the elimination of the air easier at higher temperatures.

The preheater situated at the end of the gas stream has the purpose of lowering the temperature of the gas leaving the HRSG. If the gases leave the HRSG at a high temperature we are wasting energy. Instead of wasting this energy, we can use it for heating the feed water and helping the deaerator in eliminating the gases contained in the water. In the preheater there is a recirculation circuit for rising the water temperature. That is because the water entering the heat exchanger can be at low temperature (around 30◦C) which would lead to condensation of the ﬂue gas and dew point corrosion.

The Figure 6.1 describes the combined cycle power plant with the single pressure heat recovery steam generator. The values of the heat transfer capacity (kA) are written on top of every heat exchanger. The Pinch Point in the evaporator is 8◦C and the Approach Temper- ature is 10◦C. The Figure 6.2 shows the q, T diagram of the HRSG. We have to remark that normally the reheat would be split.

The condenser is water cooled and it works at 0.04 bar of pressure and 15◦C (temperature of the water). The steam turbine delivers 117.5 MW which by taking into account the electric eﬃciency of the generator are equivalent to 112.8 MWe. These data are presented in Table 12.

CCPP Single Pressure Gas Turbine Power (MWe) 260 Steam Turbine Power (MWe) 117.5 High Pressure (bar) 120 Reheat Pressure (bar) 20 Low Pressure (bar) 4.2

Table 12: Powers and Pressures for the CCPP Single Pressure

6 Modeling a Single Pressure CCPP

133.437 133.437 77.345 77.345

324.311 324.311 6.700 6.700 1.007 bar 190.024 °C

0.4MW/K 133.437 133.437 145.347 145.347

612.219 612.219 4.700 4.700 1.008 bar 245.347 °C 130.000 bar130.000 °C 147.487 M 6.700 bar 29.016 °C 0.2MW/K 1.053 bar 23.962 °C 1.103 bar 15.000 °C 829.680 78.441 1.009 bar °C 267.787 128.000 193.787 M 829.680 829.680 0.000 2378.613 78.437 1.2MW/K 128.000 128.000 193.787 0.040 28.962 4.200 bar °C 145.354 4.200 bar °C 145.380 1448.958 0.000 1.010 bar 336.432 °C G 117555.033 kW 117555.033 126.000 318.432 122453.159 kW kW 122453.159 2.1MW/K 2672.207 2672.207 78.441 H M kJ/kg kg/s

1.010 bar °C 468.793

0.004 0.004 308.656 308.656

3084.401 3084.401 4.200 4.200 126.000 328.432

0.8MW/K

78.441 78.441

GE Turbine 9FA 50Hz 9FA Turbine GE HRSG 1P Conditions ISO 514.796 P T bar °C

3500.782 3500.782 20.000 20.000 1.011 bar °C 519.796 21.000 bar °C 298.938 0.2MW/K Figure 6.1: CCPP with a Single Pressure HRSG 1.012 bar 572.040 °C 123.000 bar 123.000 °C 439.040 G

260017.690 kW 260017.690

78.441 78.441 0.2MW/K 542.000 270851.761 kW kW 270851.761

3460.975 3460.975 120.000 120.000 1514.160 1514.160 641.000 17.000 17.000 1278.077 32.681 32.681 13.926 436.334 627.074 17.002 17.002 15.000 17.002 420.710 15.155 627.074 1.013 15.000

44 6 Modeling a Single Pressure CCPP

Temperature (°C) 700

600

500

400

300

200

100

Power (kW) 0 0 50000 100000 150000 200000 250000

q,T-Diagram 1P_DEFINITIVE.ebs 01 Secondary (Heat_exchanger) 06 Secondary (Heat_exchanger_3) 11 Primary (Evaporator) 02 Secondary (Heat_exchanger_4) 07 Secondary (Heat_exchanger_5) 12 Primary (Heat_exchanger_2) 03 Secondary (Heat_exchanger_1) 08 Primary (Heat_exchanger) 13 Primary (Heat_exchanger_3) Profile: Design 04 Secondary (Evaporator) 09 Primary (Heat_exchanger_4) 14 Primary (Heat_exchanger_5) 05 Secondary (Heat_exchanger_2) 10 Primary (Heat_exchanger_1)

Figure 6.2: Heat transfer in the HRSG in a Single Pressure CCPP

All the generators in the cycles have the same electrical efficiency, 96% in nominal conditions. For pumps the isentropic efficiency is 80% in nominal conditions and it will change in partload operation, while the mechanical efficiency is 99.8% and it will be constant in all the modes of operation. The variation of efficiencies in partload for generators and pumps is showed in Table 13 and Table 14.

Q1/Q1N ET AG/ET AGN 0 0.9 0.4 0.92 0.5 0.94 0.6 0.96 0.7 0.97 0.8 0.98 0.9 0,9 1 1 1.1 1 1.2 1

Table 13: Variation of Electrical Eﬃciency in Generators in Partload

45 7 Modeling a Two Pressure CCPP

M1/M1N ET AI/ET AIN 0 0.92 0.6 0.94 0.7 0.96 0.8 0.98 1 1 1.1 0.98

Table 14: Variation of Isentropic Eﬃciency in Pumps in Partload

7 Modeling a Two Pressure CCPP

By dividing the heat transfer process in the heat recovery steam generator in two stages, we have now two evaporators, one for high pressure and other for the low pressure. The new HRSG is composed of two superheaters at high pressure and one superheater at low pressure, two reheaters at the reheat pressure and two evaporators. The low pressure evaporator has one economizer, while the high pressure evaporator has two. We also have in this cycle the preheater at the end of the gas ﬂow and the steam turbine is divided in three stages.

The feed water, after being preheated and leaving the deaerator, is divided in two pipes and pumped to the diﬀerent pressure levels. Approximately, the 13% of the water goes to one economizer and then to the low pressure evaporator. After being evaporated and converted in saturated steam is injected in the cold reheat pipe acting as cooling steam in the reheat process. The 87% of the water goes ﬁrst through two economizers and then is evaporated in the high pressure evaporator. This process needs two economizers before reaching an adequate temperature for entering the evaporator.

In the new HRSG we have the two superheaters with the water injection between them, as in the case of a single pressure CCPP. The reheat process is divided in two stages with the water injection in between. In the cold reheat pipe saturated steam at low pressure is added for cooling and between the two reheaters water is injected. As in the single pressure process we are not injecting spray water.

CCPP Two Pressures Gas Turbine Power (MWe) 260 Steam Turbine Power (MWe) 124.4 High Pressure (bar) 120 Reheat Pressure (bar) 32 Low Pressure (bar) 5.1

Table 15: Power and Pressures for the CCPP Two Pressures

7 Modeling a Two Pressure CCPP

139.465 139.465 65.259 65.259

273.980 273.980 10.100 10.100

1.004 bar 165.236 °C

120.442 120.442 139.465 139.465

0.3MW/K 8.100 8.100 506.092 506.092 1.005 bar 212.129 °C 0.6MW/K 1.005 bar 212.129 °C 0.1MW/K 122.666 84.465 10.100 29.043 1.007 bar 252.186 °C 1.1MW/K 42.117 3200.788 1010.006 1010.006 0.000 100.580 3200.788 2802.470 2802.470 11.486 1.000 10.000 1.007 bar 281.518 °C 47.100 234.149 0.950 23.962 36.000 244.186 0.1MW/K 966.300 77.793 M 1.008 bar °C 283.469 131.200 224.275 643.418 89.279 506.092 84.465 2922.087 11.486 1.1MW/K 2336.836 84.465 5.100 152.584 5.100 120.492 34.000 34.000 278.265 1.009 bar °C 336.432 0.040 0.040 28.962 1448.958 1448.958 0.000 2.1MW/K 2672.207 77.793 G 1.009 bar °C 467.717 129.200 318.539 124385.152 kW 124385.152 129567.867 kW 126.000 126.000 328.432 H M kJ/kg kg/s 0.1MW/K 3210.282 89.279 GE TurbineGE 9FA 50Hz 2P HRSG Conditions ISO 1.010 bar 482.983 °C P T bar °C 33.000 33.000 392.983 3053.126 89.279 0.3MW/K 3108.045 3108.045 77.793 5.100 5.100 294.574

1.011 bar 537.287 °C 34.000 350.388

537.955 537.955 89.279 89.279

0.4MW/K 32.000 32.000 G 3540.486 260017.690 kW 260017.690 3191.489 3191.489 77.793 Figure 7.1: CCPP with a Two Pressure HRSG 1.012 bar °C 576.489 123.000 445.550

270851.761 kW 533.900 533.900 77.793 77.793

0.2MW/K 120.000 120.000 3439.887 3439.887 1514.160 1514.160 641.000 17.000 1278.077 32.681 32.681 13.926 436.334 627.074 17.002 17.002 15.000 17.002 420.710 15.155 15.155 627.074 1.013 1.013 15.000

47 8 Modeling a Three Pressure CCPP

Temperature (°C) 700

600

500

400

300

200

100

Power (kW) 0 0 50000 100000 150000 200000 250000 300000

q,T-Diagram 2P_DEFINITIVE.ebs 01 Secondary (Heat_exchanger) 06 Secondary (Heat_exchanger_2) 11 Secondary (Heat_exchanger_8) 02 Secondary (Heat_exchanger_5) 07 Secondary (Heat_exchanger_6) 12 Primary (Heat_exchanger) 03 Secondary (Heat_exchanger_4) 08 Secondary (Evaporator_1) 13 Primary (Heat_exchanger_5) Profile: Design 04 Secondary (Heat_exchanger_1) 09 Secondary (Heat_exchanger_7) 14 Primary (Heat_exchanger_4) 05 Secondary (Evaporator) 10 Secondary (Heat_exchanger_3) 15 Primary (Heat_exchanger_1)

Figure 7.2: Heat transfer in the HRSG with Two Pressure CCPP

8 Modeling a Three Pressure CCPP

If we add a third evaporator, the HRSG is divided now in three stages and the steam is produced now at three diﬀerent levels of pressure. The HRSG includes two superheaters at high pressure and one at the reheat pressure which superheats the steam produced in the intermediate evaporator. The high pressure evporator has two economizers, the intermediate pressure evaporator has only one while the low pressure does not have any. The steam after the low pressure evaporator is sent to the deaerator.

In the three pressure model, the water after leaving the deaerator is divided into three pipes. One part of the water (about the 15%) is evaporated at low pressure and comes back to the deaerator. Its energy is helping to heat the water in the evaporator and to remove the air contained. Other part of the water (about the 18%) goes to the intermediate pressure evaporator and is ﬁnally injected as saturated steam in the cold reheat pipe. The biggest part of the water (the left 67%) is heated ﬁrst in the econmizers and evaporated in the high pressure evaporator.

As in the two pressures cycle, we have here two superheaters with spray water injection and two reheaters with spray water injection and saturated steam injected in the cold reheat pipe.

48 8 Modeling a Three Pressure CCPP

In our three pressures model appears as well an extra circuit designed for injecting solar boosting in the CCPP. This circuit is not in operation in this model and its performance will be explained later.

8.1 Performance of Three Pressure CCPP in ISO Conditions Table 16 presents the results obtained for the Three Pressure CCPP working at the conditions for which it was designed.

CCPP Three Pressures Gas Turbine Power (MWe) 260 Steam Turbine Power (MWe) 138.5 High Pressure (bar) 155 Reheat Pressure (bar) 32 Low Pressure (bar) 4.7

Table 16: Power and Pressures for the CCPP Three Pressures

Comparing the results for the ﬁnal electric power delivered to the net, the three pressure cycle produces more power than the other two cycles. This is a consequence of having a bigger HRSG which transfers more power to the steam cycle generating more steam and increasing the eﬃciency of the cycle.

8.2 Performance of Three Pressure CCPP with Changing Ambient Temperature As we have explained for the gas turbine, its oﬄoad performance changes when the ambient temperature changes. The steam-water cycle depends on the exhaust gases of the gas turbine so its performance will change as well.

We will study the same cases studied for the gas turbine, at three diﬀerent temperatures: 0◦C, 30◦C and 45◦C.

Working in an ambient temperature lower than the design temperature causes an increment of eﬃciency of the gas turbine due to the higher density of the air which enables accepting more air into the compressor. However, the exhaust temperature of the gases is lower and with it the energy available in the gas stream for heating the water/steam in the HRSG. That means that although the gas turbine power output increases, the steam turbine power output decreases.

All the settings of the components in the steam/water cycle remain constant except in

8 Modeling a Three Pressure CCPP

347.862 °C °C 347.862 0.000 kg/s kg/s 0.000

161.000 bar bar 161.000 2577.625 kJ/kg kJ/kg 2577.625

62.727 °C °C 62.727 143.362 kg/s kg/s 143.362 29.052 °C °C 29.052 103.362 kg/s kg/s 103.362

11.200 bar bar 11.200 263.480 kJ/kg kJ/kg 263.480 11.200 bar bar 11.200 122.804 kJ/kg kJ/kg 122.804 2456.221 kJ/kg

1.003 bar bar 1.003 °C 84.985

148.717 °C °C 148.717 103.362 kg/s kg/s 103.362

3.1MW/K 4.700 bar bar 4.700 626.722 kJ/kg kJ/kg 626.722 1.004 bar 1.004 °C 161.320

1.3MW/K

149.528 °C °C 149.528 103.523 kg/s kg/s 103.523

5.200 bar bar 5.200 630.245 kJ/kg kJ/kg 630.245 1.004 bar 1.004 °C 204.055 0.1MW/K 1.005 bar 1.005 °C 206.045 5.200 bar bar 5.200 °C 149.528 122.804 kJ/kg kJ/kg 122.804 kg/s 103.362 100.580 kJ/kg kJ/kg 100.580 kg/s 6069.128 1.0MW/K 1.006 bar bar 1.006 °C 243.273 1.000 bar bar 1.000 °C 15.000 11.200 bar bar 11.200 °C 29.052 0.950 bar bar 0.950 °C 23.962 1.007 bar bar 1.007 °C 252.186 1010.006 kJ/kg kJ/kg 1010.006 kg/s 0.000 26.1MW/K M 2802.470 kJ/kg kJ/kg 2802.470 kg/s 16.647 1.3MW/K 2323.440 kJ/kg kJ/kg 2323.440 kg/s 103.362 630.177 kJ/kg kJ/kg 630.177 kg/s 103.523 H M kJ/kg kg/s 44.000 bar bar 44.000 °C 234.160 1.007 bar 1.007 °C 294.634 2577.625 kJ/kg kJ/kg 2577.625 kg/s 0.000 2848.995 kJ/kg kJ/kg 2848.995 kg/s 13.901 0.1MW/K 0.1MW/K 36.000 bar 36.000 °C 244.186 0.040 bar bar 0.040 °C 28.962 4.700 bar bar 4.700 °C 149.519 626.722 kJ/kg kJ/kg 626.722 kg/s 103.362 P T bar °C 1.008 bar 1.008 °C 298.099 166.700 bar 166.700 °C 233.009 161.000 bar bar 161.000 °C 347.862 4.700 bar 4.700 °C 196.045 1.1MW/K 1570.624 kJ/kg kJ/kg 1570.624 kg/s 0.000 4.700 bar 4.700 °C 148.717 1.009 bar 1.009 °C 355.862 G 144374.042 kW 138599.081 kW 1.9MW/K 34.000 bar 34.000 °C 288.099 164.700 bar 164.700 °C 338.066 161.000 bar 161.000 °C 347.862 1.009 bar bar 1.009 °C 457.137 0.3MW/K CC 3P-HRSG CC 50Hz 9FA GT type: ISO Conditions 1.010 bar 1.010 °C 493.337

3036.153 kJ/kg kJ/kg 3036.153 kg/s 72.975 kJ/kg 3039.401 kg/s 89.622 443.194 °C °C 443.194

0.3MW/K 158.000 bar bar 158.000 Figure 8.1: CCPP with Three Pressure HRSG at ISO Conditions 1.011 bar bar 1.011 °C 546.428 34.000 bar bar 34.000 °C 321.145 bar 4.700 °C 287.410

G 540.255 °C °C 540.255

667.303 kJ/kg kJ/kg 667.303 kg/s 641.000 259796.323 kW bar 32.000 1.012 bar bar 1.012 °C 573.535

1.013 bar bar 1.013 °C 602.296

270621.170 kW

540.296 °C °C 540.296 0.2MW/K 0.3MW/K kg/s 72.975

155.000 bar bar 155.000 3418.506 kJ/kg kJ/kg 3418.506 1514.784 kJ/kg kJ/kg 1514.784 kg/s 641.000 15.155 kJ/kg kJ/kg 15.155 kg/s 627.074 17.000 bar 17.000 °C 1278.554 1.010 bar 1.010 °C 15.000

50 8 Modeling a Three Pressure CCPP

Temperature (°C) 700

600

500

400

300

200

100

Power (kW) 0 0 50000 100000 150000 200000 250000 300000 350000

q,T-Diagram 3P_SOLAR_v2_difftemp.ebs 01 Secondary (Heat_exchanger) 06 Secondary (Heat_exchanger_2) 11 Secondary (Heat_exchanger_9) 02 Secondary (Heat_exchanger_5) 07 Secondary (Heat_exchanger_6) 12 Secondary (Evaporator_2) 03 Secondary (Heat_exchanger_4) 08 Secondary (Evaporator_1) 13 Secondary (Heat_exchanger_8) Profile: Design 04 Secondary (Heat_exchanger_1) 09 Secondary (Heat_exchanger_7) 14 Primary (Heat_exchanger) 05 Secondary (Evaporator) 10 Secondary (Heat_exchanger_3) 15 Primary (Heat_exchanger_5)

Figure 8.2: Heat transfer in the HRSG with Three Pressure CCPP at ISO Conditions the condenser, where the cooling water suﬀers a variation in temperature at the same time that the ambient temperature changes. In this case, when the ambient temperature is 0◦C, we suppose the sea water has a temperature of 10◦C.

Cases Gas Turbine Power Steam Turbine Power m˙ steam (kg/s) m˙ gas (kg/s) ISO 260 MWe 138.5 MWe 72.9 641 ◦ 0 C 318.6 MWe 134.6 MWe 70.4 677 ◦ 30 C 213.3 MWe 122.5 MWe 69.2 608.5 ◦ 45 C 183.6 MWe 115.4 MWe 66.9 588.4

Table 17: Comparison between the Partload Performance and Design Performance

The Table 17 shows the power and mass ﬂuctuation when the ambient temperature changes. As we can see, although the gas turbine delivers more power when the ambient temperature is lower, the steam turbine always has its maximum power at design conditions. The closer to ISO conditions the higher power delivered.

8 Modeling a Three Pressure CCPP

343.850 °C °C 343.850 0.000 kg/s kg/s 0.000

153.198 bar bar 153.198 2601.582 kJ/kg kJ/kg 2601.582

86.052 °C °C 86.052 222.934 kg/s kg/s 222.934 23.736 °C °C 23.736 101.934 kg/s kg/s 101.934

13.753 bar bar 13.753 361.395 kJ/kg kJ/kg 361.395 13.753 bar bar 13.753 100.827 kJ/kg kJ/kg 100.827 2502.476 kJ/kg

1.002 bar bar 1.002 °C 94.436

137.894 °C °C 137.894 101.934 kg/s kg/s 101.934

3.3MW/K 4.581 bar bar 4.581 580.241 kJ/kg kJ/kg 580.241 1.003 bar 1.003 °C 161.942

1.3MW/K

148.577 °C °C 148.577 104.033 kg/s kg/s 104.033

5.184 bar bar 5.184 626.147 kJ/kg kJ/kg 626.147 1.003 bar 1.003 °C 206.364 0.1MW/K 1.004 bar 1.004 °C 208.485 5.184 bar bar 5.184 °C 148.577 100.827 kJ/kg kJ/kg 100.827 kg/s 101.934 78.745 kJ/kg kJ/kg 78.745 kg/s 6066.211 1.0MW/K C 1.005 bar bar 1.005 °C 243.279 1.000 bar bar 1.000 °C 10.000 ◦ 13.753 bar bar 13.753 °C 23.736 0.950 bar bar 0.950 °C 18.743 1.006 bar bar 1.006 °C 252.440 1002.037 kJ/kg kJ/kg 1002.037 kg/s 0.000 26.1MW/K M 2802.565 kJ/kg kJ/kg 2802.565 kg/s 18.323 1.3MW/K 2278.866 kJ/kg kJ/kg 2278.866 kg/s 101.934 626.065 kJ/kg kJ/kg 626.065 kg/s 104.033 H M kJ/kg kg/s 45.316 bar bar 45.316 °C 232.458 1.006 bar 1.006 °C 296.754 2601.582 kJ/kg kJ/kg 2601.582 kg/s 0.000 2851.599 kJ/kg kJ/kg 2851.599 kg/s 15.274 0.1MW/K 0.1MW/K 35.674 bar 35.674 °C 243.661 0.029 bar bar 0.029 °C 23.629 4.581 bar bar 4.581 °C 148.567 580.241 kJ/kg kJ/kg 580.241 kg/s 101.934 P T bar °C 1.007 bar 1.007 °C 300.470 158.507 bar 158.507 °C 233.891 153.198 bar bar 153.198 °C 343.850 4.581 bar 4.581 °C 196.970 1.1MW/K 1561.904 kJ/kg kJ/kg 1561.904 kg/s 0.000 4.581 bar 4.581 °C 137.894 1.009 bar 1.009 °C 351.959 G 141330.765 kW 134619.167 kW 2.0MW/K 33.256 bar 33.256 °C 288.888 156.640 bar 156.640 °C 336.442 153.198 bar 153.198 °C 343.850 1.009 bar bar 1.009 °C 447.425 Figure 8.3: Three Pressure CCPP at 0 0.3MW/K CC 3P-HRSG CC GT type: 9FA 50Hz Atmospheric Temperature: 0 ºC 1.010 bar 1.010 °C 481.321

3007.447 kJ/kg kJ/kg 3007.447 kg/s 70.436 kJ/kg 3008.321 kg/s 88.759 435.130 °C °C 435.130

0.3MW/K 150.397 bar bar 150.397 1.011 bar bar 1.011 °C 526.588 33.256 bar bar 33.256 °C 309.018 bar 4.581 °C 272.165

G 520.139 °C °C 520.139

634.289 kJ/kg kJ/kg 634.289 kg/s 677.018 318634.209 kW bar 31.292 1.012 bar bar 1.012 °C 549.747

1.013 bar bar 1.013 °C 573.043

331910.634 kW

520.046 °C °C 520.046 0.2MW/K 0.3MW/K kg/s 70.436

147.597 bar bar 147.597 3370.819 kJ/kg kJ/kg 3370.819 1518.409 kJ/kg kJ/kg 1518.409 kg/s 677.018 -7.244 kJ/kg kJ/kg -7.244 kg/s 661.528 17.955 bar 17.955 °C 1278.077 1.013 bar 1.013 °C 0.000

8 Modeling a Three Pressure CCPP

344.110 °C °C 344.110 0.000 kg/s kg/s 0.000

153.693 bar bar 153.693 2600.118 kJ/kg kJ/kg 2600.118

86.458 °C °C 86.458 182.538 kg/s kg/s 182.538 38.357 °C °C 38.357 96.538 kg/s kg/s 96.538

11.544 bar bar 11.544 362.927 kJ/kg kJ/kg 362.927 11.544 bar bar 11.544 161.696 kJ/kg kJ/kg 161.696 2439.876 kJ/kg

1.004 bar bar 1.004 °C 94.527

139.794 °C °C 139.794 96.538 kg/s kg/s 96.538

3.3MW/K 4.406 bar bar 4.406 588.368 kJ/kg kJ/kg 588.368 1.005 bar 1.005 °C 158.202

1.2MW/K

147.137 °C °C 147.137 97.904 kg/s kg/s 97.904

4.843 bar bar 4.843 619.927 kJ/kg kJ/kg 619.927 1.005 bar 1.005 °C 200.623 0.1MW/K 1.006 bar 1.006 °C 202.576 4.843 bar bar 4.843 °C 147.137 161.696 kJ/kg kJ/kg 161.696 kg/s 96.538 140.511 kJ/kg kJ/kg 140.511 kg/s 6066.211 1.0MW/K C 1.007 bar bar 1.007 °C 239.615 ◦ 1.000 bar bar 1.000 °C 25.000 11.544 bar bar 11.544 °C 38.357 0.950 bar bar 0.950 °C 33.513 1.008 bar bar 1.008 °C 248.449 998.279 kJ/kg kJ/kg 998.279 kg/s 0.000 26.1MW/K M 2802.973 kJ/kg kJ/kg 2802.973 kg/s 15.652 1.3MW/K 2396.167 kJ/kg kJ/kg 2396.167 kg/s 96.538 619.868 kJ/kg kJ/kg 619.868 kg/s 97.904 H M kJ/kg kg/s 40.988 bar 40.988 °C 231.672 1.008 bar 1.008 °C 290.931 2600.118 kJ/kg kJ/kg 2600.118 kg/s 0.000 2845.005 kJ/kg kJ/kg 2845.005 kg/s 13.013 0.1MW/K 0.1MW/K 33.958 bar 33.958 °C 240.830 0.067 bar bar 0.067 °C 38.254 4.406 bar bar 4.406 °C 147.130 588.368 kJ/kg kJ/kg 588.368 kg/s 96.538 P T bar °C 1.008 bar 1.008 °C 294.340 158.800 bar 158.800 °C 229.963 153.693 bar bar 153.693 °C 344.110 4.406 bar 4.406 °C 193.438 1.1MW/K 1553.026 kJ/kg kJ/kg 1553.026 kg/s 0.000 4.406 bar 4.406 °C 139.794 1.009 bar 1.009 °C 351.937 G 130111.265 kW 122494.157 kW 1.8MW/K 32.194 bar 32.194 °C 285.334 156.996 bar 156.996 °C 335.241 153.693 bar 153.693 °C 344.110 1.009 bar bar 1.009 °C 457.544 Figure 8.4: Three Pressure CCPP at 30 0.3MW/K CC 3P-HRSG 50Hz 9FA type: GT ºC 30 Temperature: Atmospheric 1.010 bar 1.010 °C 493.410

3052.175 kJ/kg kJ/kg 3052.175 kg/s 69.240 kJ/kg 3044.636 kg/s 84.891 448.883 °C °C 448.883

0.3MW/K 150.987 bar bar 150.987 1.011 bar bar 1.011 °C 547.592 32.194 bar bar 32.194 °C 325.629 bar 4.406 °C 289.568

G 543.041 °C °C 543.041

665.491 kJ/kg kJ/kg 665.491 kg/s 608.469 213291.934 kW bar 30.397 1.012 bar bar 1.012 °C 574.456

1.013 bar bar 1.013 °C 602.302

228553.204 kW

545.693 °C °C 545.693 0.2MW/K 0.3MW/K kg/s 69.240

148.281 bar bar 148.281 3440.620 kJ/kg kJ/kg 3440.620 1478.070 kJ/kg kJ/kg 1478.070 kg/s 608.469 30.317 kJ/kg kJ/kg 30.317 kg/s 596.030 16.011 bar 16.011 °C 1253.964 1.010 bar 1.010 °C 30.000

8 Modeling a Three Pressure CCPP

341.548 °C °C 341.548 0.000 kg/s kg/s 0.000

148.861 bar bar 148.861 2614.102 kJ/kg kJ/kg 2614.102

74.030 °C °C 74.030 134.296 kg/s kg/s 134.296 43.123 °C °C 43.123 94.296 kg/s kg/s 94.296

9.801 bar bar 9.801 310.670 kJ/kg kJ/kg 310.670 9.801 bar bar 9.801 181.444 kJ/kg kJ/kg 181.444 2433.893 kJ/kg

1.005 bar bar 1.005 °C 91.245

146.015 °C °C 146.015 94.296 kg/s kg/s 94.296

3.0MW/K 4.310 bar bar 4.310 615.067 kJ/kg kJ/kg 615.067 1.005 bar 1.005 °C 156.817

1.2MW/K

146.334 °C °C 146.334 94.353 kg/s kg/s 94.353

4.708 bar bar 4.708 616.463 kJ/kg kJ/kg 616.463 1.005 bar 1.005 °C 198.742 0.1MW/K 1.006 bar 1.006 °C 200.650 4.708 bar bar 4.708 °C 146.334 181.444 kJ/kg kJ/kg 181.444 kg/s 94.296 160.814 kJ/kg kJ/kg 160.814 kg/s 6069.128 0.9MW/K C 1.007 bar bar 1.007 °C 237.316 ◦ 1.000 bar bar 1.000 °C 30.000 9.801 bar bar 9.801 °C 43.123 0.950 bar bar 0.950 °C 38.371 1.008 bar bar 1.008 °C 246.027 991.095 kJ/kg kJ/kg 991.095 kg/s 0.000 26.1MW/K M 2803.171 kJ/kg kJ/kg 2803.171 kg/s 15.022 1.2MW/K 2431.687 kJ/kg kJ/kg 2431.687 kg/s 94.296 616.409 kJ/kg kJ/kg 616.409 kg/s 94.353 H M kJ/kg kg/s 39.153 bar 39.153 °C 230.144 1.008 bar 1.008 °C 288.482 2614.102 kJ/kg kJ/kg 2614.102 kg/s 0.000 2842.707 kJ/kg kJ/kg 2842.707 kg/s 12.393 0.1MW/K 0.1MW/K 32.679 bar 32.679 °C 238.650 0.087 bar bar 0.087 °C 43.031 4.310 bar bar 4.310 °C 146.327 615.067 kJ/kg kJ/kg 615.067 kg/s 94.296 P T bar °C 1.009 bar 1.009 °C 291.847 153.603 bar 153.603 °C 228.138 148.861 bar bar 148.861 °C 341.548 4.310 bar 4.310 °C 192.115 1.1MW/K 1541.051 kJ/kg kJ/kg 1541.051 kg/s 0.000 4.310 bar 4.310 °C 146.015 1.010 bar 1.010 °C 349.237 G 123436.150 kW 115388.206 kW 1.8MW/K 31.051 bar 31.051 °C 283.444 151.920 bar 151.920 °C 333.310 148.861 bar 148.861 °C 341.548 1.010 bar bar 1.010 °C 457.725 Figure 8.5: Three Pressure CCPP at 45 0.3MW/K CC 3P-HRSG 50Hz 9FA type: GT ºC 45 Temperature: Atmospheric 1.010 bar 1.010 °C 493.322

3062.655 kJ/kg kJ/kg 3062.655 kg/s 66.938 kJ/kg 3052.441 kg/s 81.959 452.784 °C °C 452.784

0.3MW/K 146.337 bar bar 146.337 1.011 bar bar 1.011 °C 548.304 31.051 bar bar 31.051 °C 328.702 bar 4.310 °C 293.242

G 544.898 °C °C 544.898

664.024 kJ/kg kJ/kg 664.024 kg/s 588.395 183627.727 kW bar 29.378 1.012 bar bar 1.012 °C 575.083

1.013 bar bar 1.013 °C 602.330

199929.198 kW

549.164 °C °C 549.164 0.2MW/K 0.3MW/K kg/s 66.938

143.812 bar bar 143.812 3454.749 kJ/kg kJ/kg 3454.749 1455.053 kJ/kg kJ/kg 1455.053 kg/s 588.395 45.488 kJ/kg kJ/kg 45.488 kg/s 576.988 15.405 bar 15.405 °C 1239.115 1.010 bar 1.010 °C 45.000

54 9 Standard Three Pressure CCPP with Solar Boosting

9 Standard Three Pressure CCPP with Solar Boosting

The eﬃciency of the Rankine cycle increases with the steam pressure and the amount of steam expanding through the turbine. When the ambient temperature rises the mass ﬂow expanded through the gas turbine decreases and the energy available in the gas stream decreases as well. For this reason the heat transferred in the HRSG is smaller. That generates a smaller amount of steam in the HRSG and the power output of the steam turbine decreases due to the less mass of steam expanded.

At hot temperature conditions the overall efficiency of the CCPP decreases. Usually hot temperatures and sunny sky come together so if the location of our CCPP is in a place where the incident radiation of the Sun is enough for running a solar field, the energy which that field provides can help the electricity production in the CCPP. With a parabolic trough system installed next to the combined cycle plant, the incident radiation can be used as heat energy into the Rankine cycle of the CCPP. With this energy the lack of input energy in the Rankine cycle is compensated and the power output of the steam cycle increases.

That configuration, called Solar Boosting, is only possible when the Sun is shining and the mass flow of steam going through the steam turbine is smaller than in design conditions. We consider a parabolic trough solar field next to the CCPP in which the steam is generated directly using a once through system. In the solar field the water is converted into saturated steam and injected in the Rankine cycle. In state of art technology the solar field would be cooled by synthetic oil and the steam would be generated in an additional evaporator, see Figure 3.7. In this work we analyze an advanced system based on direct steam generation.

For modeling the solar boosting we have designed an extra circuit, shown in Figure 9.1. The water, after leaving the condenser, is pumped to the drum of the forced circulation solar steam generator. The water evaporates into saturated steam and goes through the steam drum. There, water in the mixture it is separated from the steam and recirculated in the solar boosting cycle. The concentrator heat exchangers of the solar ﬁeld are not simulated in Ebsilon. Instead we use a ”boundary value component” for setting the solar heat added in each case. We set a recirculation mass ﬂow in the solar boosting circuit always higher than ten times the mass of saturated steam generated.

The amount of heat added is calculated as the heat necessary to heat the water from the conditions after leaving the condenser until saturated steam at high pressure. This specific heat is the difference between the enthalpy of the condensed water at high pressure and the enthalpy of the steam generated in the high pressure evaporator. The specific heat is multiplied by the mass flow added in each case getting the final value of heat added to the cycle.

There are four diﬀerent possibilities for injecting the steam into the Rankine cycle depending on the point of injection and the properties of the injected steam: • Cold Reheated Steam: the steam is injected after the ﬁrst expansion in the turbine before the reheat exchanger. • Hot Reheated Steam: the steam is injected after the reheat, just before entering in the second stage of the turbine.

55 9 Standard Three Pressure CCPP with Solar Boosting

Solar Heat Injection

Figure 9.1: Solar Boosting Components

• Saturated High Pressure Steam: the solar ﬁeld generates saturated steam at high pressure which is injected after the high pressure evaporator.

• Hot High Pressure Steam: in this conﬁguration the steam is injected at high pressure in the hottest point of the circuit, after the superheaters.

The Figures 9.2 and 9.3 illustrate these deﬁnitions.

In this work we only discuss the Cold Reheated (CRH) and the Saturated High Pressure (SHP) steam configuration because they are the robust configurations. Systems with Hot Reheated (HRH) steam or Hot High Pressure (HHP) steam may have problems of thermal stress due to the temperature difference of the fluids in the injection point. In addition, solar absorbers for 540◦C steam temperature are not yet available and the number of solar boosting hours at high pressure is expected to be lower.

9.1 Standard Three Pressure CCPP with Solar Boosting and Saturated High Pressure Steam (SHP) In the Standard Study of Solar Boosting, saturated steam at high pressure will be injected in the HRSG in a point of the pipe which connects the HP evaporator with the superheater. The properties of the injected steam depend on the case of study.

The case of study is that in which the temperature is warmer than in design conditions. It is assumed that the incident radiation is able to heat the receiver fluid in the parabolic trough. Although at ISO conditions (15◦C) it is possible to obtain heat energy from the collector field we suppose at these conditions the CCPP is working at full operation and the solar heat is not necessary. We suppose that at 0◦C, in winter, the Sun hardly shines and it is not possible to obtain energy from the solar field. The only considered cases will be 30◦C and 45◦C.

APPROACH ”A”

56 9 Standard Three Pressure CCPP with Solar Boosting

Saturated High Pressure Steam High Pressure High Pressure High Pressure Superheater Reheater Superheater Reheater Evaporator

Cold Reheated Steam

Figure 9.2: Diagram of Cold Reheated Steam and Saturated High Pressure Steam injection points.

With the Solar Boosting study we pretend to ﬁnd out the maximum solar heat which can be injected into an existing Rankine cycle. That means that the design parameters must remain unalterable:

• the pressure in the steam drum can not exceed the design pressure in nominal conditions.

• the HRSG shall not be modiﬁed.

• the high pressure and reheat temperatures must remain constant.

• the amount of solar heat added is as big as possible but the combination of solar steam and the HRSG steam is never bigger than the HRSG steam in design conditions.

In every case we considered different amounts of solar heat added. The three chosen loading cases are: adding a 5%, 10% or 12% of the nominal steam flow. It is not possible to add more than 12% of nominal steam because the section of the steam turbine does not accept more flow.

The steam mass ﬂow at ISO conditions are 72.9 kg/s and the corresponding masses added in each case are presented in Table 18.

57 9 Standard Three Pressure CCPP with Solar Boosting

High Pressure High Pressure High Pressure Superheater Reheater Superheater Reheater Evaporator

Hot High Pressure Steam

Hot Reheated Steam

Figure 9.3: Diagram of Hot Reheated Steam and Hot High Pressure Steam injection points.

m˙ solar (kg/s) 5% 3.6 10% 7.3 12% 8.7

Table 18: Mass ﬂow of Steam added in each case.

Looking at the results shown in the Table 19 we can conclude that the addition of steam coming from the solar ﬁeld always increases the steam cycle power output because more steam is being expanded in the steam turbine.

The table 19 also compares the results obtained for each cycle when solar heat is added. As we can see in the table the bigger amount of solar heat added the bigger output power of the steam turbine. Besides, by adding the same percentage of steam heated by solar energy, the output power is bigger at lower temperatures. That is because the ﬂuegas mass ﬂow of the gas turbine is higher at lower temperatures, so the result is a higher heat input to the steam cycle because more energy is being transferred in the HRSG. The maximum power that we can obtain in our study is at 30◦C by adding the 12% of steam in design case.

The models of the three cases at 30◦C are shown in Figures 9.4, 9.5 and ??nd at 45◦C in the Figures 9.7, 9.8 and 9.9.

58 9 Standard Three Pressure CCPP with Solar Boosting

Cases Psteamturbine (MWe) m˙ HP EV AP (kg/s) Qsolar(MW) ISO 138.48 72.9 NO 0◦C 134.62 70.4 SOLAR 30◦C 122.49 69.2 45◦C 115.4 66.9 SOLAR, 5% of nominal 30◦C 125.84 67.43 8.85 steam flow (3.6 kg/s) 45◦C 118.58 65.12 8.83 SOLAR, 10% of nominal 30◦C 129.16 65.6 17.61 steam flow (7.3 kg/s) 45◦C 121,8 63.3 17.57 SOLAR, 12% of nominal 30◦C 130.36 64.9 21.09 steam flow (8.7 kg/s) 45◦C 123.03 62.57 21.04

Table 19: Comparison between Three Pressure Cycles when Solar Heat is added.

Looking at Figure 9.6 we can see that compared to the operation without solar boosting, see Figure 8.4, the high pressure rises by 7 bar and the high pressure and reheat temperatures drop by 7-9 ◦C. After seeing the results of this study two interesting questions have to be solved:

1. Is the solar ﬁeld construction proﬁtable?

2. It is economic to built the CCPP in a way to use solar steam at nominal conditions?

The discussion of these two question will be done in the consecutive studies in chapters 9 and 10.

9.2 Standard Three Pressure CCPP with Solar Boosting and Cold Reheated Steam (CRH) Other possibility for injecting the steam is the Cold Reheated (CRH) steam injection. As we explain before, with this conﬁguration the steam heated with solar energy is added to the cycle after the ﬁrst expansion stage in the steam turbine.

We will divide this study in two parts:

• In the ﬁrst part, the amount of energy injected in the cycle will be exactly the same than in the case of injecting saturated high pressure (SHP) steam. With this we want to compare the two methods.

• In the second part, the amount of energy injected is the energy needed for increasing the temperature of the feed water to the cold reheat temperature after the ﬁrst expansion stage in the steam turbine. The amount of energy is calculated by multiplying the enthalpy diﬀerence between the point of injection and the water after the condenser and the mass of steam heated.

The results of the ﬁrst part are shown in the Table 20. Comparing the output power of the steam turbine in the cases of CRH and SHP we can conclude that by adding the same

9 Standard Three Pressure CCPP with Solar Boosting

345.869 °C °C 345.869 3.644 kg/s kg/s 3.644

157.084 bar bar 157.084 2589.890 kJ/kg kJ/kg 2589.890

87.529 °C °C 87.529 181.421 kg/s kg/s 181.421 38.655 °C °C 38.655 95.421 kg/s kg/s 95.421

11.517 bar bar 11.517 367.425 kJ/kg kJ/kg 367.425 11.517 bar bar 11.517 162.940 kJ/kg kJ/kg 162.940 2428.402 kJ/kg

1.004 bar bar 1.004 °C 95.707

141.075 °C °C 141.075 95.421 kg/s kg/s 95.421

3.3MW/K 4.508 bar bar 4.508 593.868 kJ/kg kJ/kg 593.868 1.005 bar 1.005 °C 159.262

1.2MW/K

147.981 °C °C 147.981 96.692 kg/s kg/s 96.692

4.964 bar bar 4.964 623.568 kJ/kg kJ/kg 623.568 1.005 bar 1.005 °C 202.470 0.1MW/K 1.006 bar 1.006 °C 204.494 4.964 bar bar 4.964 °C 147.981 162.940 kJ/kg kJ/kg 162.940 kg/s 95.421 141.311 kJ/kg kJ/kg 141.311 kg/s 6069.128 0.9MW/K 1.007 bar bar 1.007 °C 240.964 1.000 bar bar 1.000 °C 25.000 11.517 bar bar 11.517 °C 38.655 0.950 bar bar 0.950 °C 33.704 1.008 bar bar 1.008 °C 250.009 1003.217 kJ/kg kJ/kg 1003.217 kg/s 0.000 26.1MW/K M 2802.789 kJ/kg kJ/kg 2802.789 kg/s 15.979 1.3MW/K 2390.446 kJ/kg kJ/kg 2390.446 kg/s 99.065 623.505 kJ/kg kJ/kg 623.505 kg/s 96.692 H M kJ/kg kg/s 42.170 bar bar 42.170 °C 232.721 1.008 bar 1.008 °C 293.239 2589.890 kJ/kg kJ/kg 2589.890 kg/s 3.644 2847.706 kJ/kg kJ/kg 2847.706 kg/s 13.274 0.1MW/K 0.1MW/K 34.815 bar 34.815 °C 242.257 0.068 bar bar 0.068 °C 38.552 4.508 bar bar 4.508 °C 147.973 593.868 kJ/kg kJ/kg 593.868 kg/s 95.421 P T bar °C 1.008 bar 1.008 °C 296.782 161.946 bar 161.946 °C 231.712 157.084 bar bar 157.084 °C 345.869 4.508 bar 4.508 °C 194.962 1.1MW/K 1565.814 kJ/kg kJ/kg 1565.814 kg/s 0.000 4.508 bar 4.508 °C 141.075 C with a 5% of high pressure steam added as Solar Heat 1.009 bar 1.009 °C 353.325 ◦ G 133239.167 kW 125833.440 kW 1.8MW/K 32.972 bar 32.972 °C 287.295 160.238 bar 160.238 °C 337.169 157.084 bar 157.084 °C 345.869 1.009 bar bar 1.009 °C 453.952 0.3MW/K CC 3P-HRSG CC GT 9FA 50Hz type: ºC 30 Temperature: Atmospheric Injection Heat Solar 5% 1.010 bar 1.010 °C 490.409

3042.190 kJ/kg kJ/kg 3042.190 kg/s 71.083 kJ/kg 3038.596 kg/s 87.062 444.326 °C °C 444.326

0.3MW/K 154.237 bar bar 154.237 1.011 bar bar 1.011 °C 544.958 32.972 bar bar 32.972 °C 322.437 bar 4.508 °C 286.766

G 540.144 °C °C 540.144

665.505 kJ/kg kJ/kg 665.505 kg/s 608.469 213307.159 kW bar 31.084 1.012 bar bar 1.012 °C 572.920

1.013 bar bar 1.013 °C 602.313

228545.622 kW

542.099 °C °C 542.099 0.2MW/K 0.3MW/K kg/s 71.083

151.391 bar bar 151.391 3427.445 kJ/kg kJ/kg 3427.445 1478.034 kJ/kg kJ/kg 1478.034 kg/s 608.469 Figure 9.4: Three Pressure CCPP at 30 30.317 kJ/kg kJ/kg 30.317 kg/s 596.030 16.009 bar 16.009 °C 1253.936 1.010 bar 1.010 °C 30.000

9 Standard Three Pressure CCPP with Solar Boosting

347.631 °C °C 347.631 7.289 kg/s kg/s 7.289

160.542 bar bar 160.542 2579.085 kJ/kg kJ/kg 2579.085

88.616 °C °C 88.616 180.304 kg/s kg/s 180.304 38.957 °C °C 38.957 94.304 kg/s kg/s 94.304

11.497 bar bar 11.497 371.987 kJ/kg kJ/kg 371.987 11.497 bar bar 11.497 164.200 kJ/kg kJ/kg 164.200 2416.337 kJ/kg

1.004 bar bar 1.004 °C 96.895

142.364 °C °C 142.364 94.304 kg/s kg/s 94.304

3.3MW/K 4.612 bar bar 4.612 599.401 kJ/kg kJ/kg 599.401 1.005 bar 1.005 °C 160.318

1.2MW/K

148.822 °C °C 148.822 95.480 kg/s kg/s 95.480

5.086 bar bar 5.086 627.195 kJ/kg kJ/kg 627.195 1.005 bar 1.005 °C 204.343 0.1MW/K 1.006 bar 1.006 °C 206.442 5.086 bar bar 5.086 °C 148.822 164.200 kJ/kg kJ/kg 164.200 kg/s 94.304 142.120 kJ/kg kJ/kg 142.120 kg/s 6069.128 0.9MW/K 1.007 bar bar 1.007 °C 242.304 1.000 bar bar 1.000 °C 25.000 11.497 bar bar 11.497 °C 38.957 0.950 bar bar 0.950 °C 33.898 1.008 bar bar 1.008 °C 251.568 1008.181 kJ/kg kJ/kg 1008.181 kg/s 0.000 26.1MW/K M 2802.560 kJ/kg kJ/kg 2802.560 kg/s 16.321 1.3MW/K 2384.611 kJ/kg kJ/kg 2384.611 kg/s 101.593 627.130 kJ/kg kJ/kg 627.130 kg/s 95.480 H M kJ/kg kg/s 43.375 bar bar 43.375 °C 233.774 1.008 bar 1.008 °C 295.575 2579.085 kJ/kg kJ/kg 2579.085 kg/s 7.289 2850.459 kJ/kg kJ/kg 2850.459 kg/s 13.545 0.1MW/K 0.1MW/K 35.690 bar 35.690 °C 243.686 0.069 bar bar 0.069 °C 38.853 4.612 bar bar 4.612 °C 148.814 599.401 kJ/kg kJ/kg 599.401 kg/s 94.304 P T bar °C 1.008 bar 1.008 °C 299.260 165.171 bar 165.171 °C 233.445 160.542 bar bar 160.542 °C 347.631 4.612 bar 4.612 °C 196.510 1.1MW/K 1578.741 kJ/kg kJ/kg 1578.741 kg/s 0.000 4.612 bar 4.612 °C 142.364 C with a 10% of high pressure steam added as Solar Heat 1.009 bar 1.009 °C 354.725 G ◦ 136312.432 kW 129148.082 kW 1.8MW/K 33.768 bar 33.768 °C 289.286 163.554 bar 163.554 °C 339.092 160.542 bar 160.542 °C 347.631 1.009 bar bar 1.009 °C 450.386 0.3MW/K CC 3P-HRSG CC GT 9FA 50Hz type: ºC 30 Temperature: Atmospheric Injection Heat Solar 10% 1.010 bar 1.010 °C 487.482

3031.577 kJ/kg kJ/kg 3031.577 kg/s 72.903 kJ/kg 3032.332 kg/s 89.224 439.743 °C °C 439.743

0.3MW/K 157.548 bar bar 157.548 1.011 bar bar 1.011 °C 542.288 33.768 bar bar 33.768 °C 319.076 bar 4.612 °C 283.862

G 537.139 °C °C 537.139

665.505 kJ/kg kJ/kg 665.505 kg/s 608.469 213307.159 kW bar 31.785 1.012 bar bar 1.012 °C 571.331

1.013 bar bar 1.013 °C 602.313

228545.622 kW

538.278 °C °C 538.278 0.2MW/K 0.3MW/K kg/s 72.903

154.554 bar bar 154.554 3413.423 kJ/kg kJ/kg 3413.423 1478.034 kJ/kg kJ/kg 1478.034 kg/s 608.469 Figure 9.5: Three Pressure CCPP at 30 30.317 kJ/kg kJ/kg 30.317 kg/s 596.030 16.009 bar 16.009 °C 1253.936 1.010 bar 1.010 °C 30.000

9 Standard Three Pressure CCPP with Solar Boosting

348.343 °C °C 348.343 8.747 kg/s kg/s 8.747

161.958 bar bar 161.958 2574.547 kJ/kg kJ/kg 2574.547

89.051 °C °C 89.051 179.879 kg/s kg/s 179.879 39.083 °C °C 39.083 93.879 kg/s kg/s 93.879

11.495 bar bar 11.495 373.817 kJ/kg kJ/kg 373.817 11.495 bar bar 11.495 164.725 kJ/kg kJ/kg 164.725 2411.274 kJ/kg

1.004 bar bar 1.004 °C 97.371

142.883 °C °C 142.883 93.879 kg/s kg/s 93.879

3.3MW/K 4.655 bar bar 4.655 601.632 kJ/kg kJ/kg 601.632 1.005 bar 1.005 °C 160.751

1.2MW/K

149.173 °C °C 149.173 95.020 kg/s kg/s 95.020

5.138 bar bar 5.138 628.711 kJ/kg kJ/kg 628.711 1.005 bar 1.005 °C 205.094 0.1MW/K 1.006 bar 1.006 °C 207.222 5.138 bar bar 5.138 °C 149.173 164.725 kJ/kg kJ/kg 164.725 kg/s 93.879 142.457 kJ/kg kJ/kg 142.457 kg/s 6069.128 0.9MW/K 1.007 bar bar 1.007 °C 242.847 1.000 bar bar 1.000 °C 25.000 11.495 bar bar 11.495 °C 39.083 0.950 bar bar 0.950 °C 33.979 1.008 bar bar 1.008 °C 252.197 1010.218 kJ/kg kJ/kg 1010.218 kg/s 0.000 26.1MW/K M 2802.455 kJ/kg kJ/kg 2802.455 kg/s 16.453 1.3MW/K 2382.689 kJ/kg kJ/kg 2382.689 kg/s 102.626 628.645 kJ/kg kJ/kg 628.645 kg/s 95.020 H M kJ/kg kg/s 43.865 bar bar 43.865 °C 234.205 1.008 bar 1.008 °C 296.499 2574.547 kJ/kg kJ/kg 2574.547 kg/s 8.747 2851.556 kJ/kg kJ/kg 2851.556 kg/s 13.651 0.1MW/K 0.1MW/K 36.049 bar 36.049 °C 244.265 0.070 bar bar 0.070 °C 38.979 4.655 bar bar 4.655 °C 149.165 601.632 kJ/kg kJ/kg 601.632 kg/s 93.879 P T bar °C 1.008 bar 1.008 °C 300.241 166.500 bar 166.500 °C 234.140 161.958 bar bar 161.958 °C 348.343 4.655 bar 4.655 °C 197.133 1.1MW/K 1583.945 kJ/kg kJ/kg 1583.945 kg/s 0.000 4.655 bar 4.655 °C 142.883 C with a 12% of high pressure steam added as Solar Heat 1.009 bar 1.009 °C 355.296 G ◦ 137435.758 kW 130364.420 kW 1.8MW/K 34.096 bar 34.096 °C 290.073 164.917 bar 164.917 °C 339.860 161.958 bar 161.958 °C 348.343 1.009 bar bar 1.009 °C 449.027 0.3MW/K CC 3P-HRSG CC GT 9FA 50Hz type: ºC 30 Temperature: Atmospheric Injection Heat Solar 12% 1.010 bar 1.010 °C 486.326

3027.997 kJ/kg kJ/kg 3027.997 kg/s 73.662 kJ/kg 3030.626 kg/s 90.116 437.826 °C °C 437.826

0.3MW/K 158.901 bar bar 158.901 1.011 bar bar 1.011 °C 541.203 34.096 bar bar 34.096 °C 318.024 bar 4.655 °C 283.093

G 535.889 °C °C 535.889

665.505 kJ/kg kJ/kg 665.505 kg/s 608.469 213307.159 kW bar 32.073 1.012 bar bar 1.012 °C 570.668

1.013 bar bar 1.013 °C 602.313

228545.622 kW

536.577 °C °C 536.577 0.2MW/K 0.3MW/K kg/s 73.662

155.844 bar bar 155.844 3407.256 kJ/kg kJ/kg 3407.256 1478.034 kJ/kg kJ/kg 1478.034 kg/s 608.469 Figure 9.6: Three Pressure CCPP at 30 30.317 kJ/kg kJ/kg 30.317 kg/s 596.030 16.009 bar 16.009 °C 1253.936 1.010 bar 1.010 °C 30.000

9 Standard Three Pressure CCPP with Solar Boosting

343.457 °C °C 343.457 3.644 kg/s kg/s 3.644

152.449 bar bar 152.449 2603.780 kJ/kg kJ/kg 2603.780

77.306 °C °C 77.306 138.042 kg/s kg/s 138.042 43.413 °C °C 43.413 93.042 kg/s kg/s 93.042

9.902 bar bar 9.902 324.406 kJ/kg kJ/kg 324.406 9.902 bar bar 9.902 182.665 kJ/kg kJ/kg 182.665 2422.365 kJ/kg

1.005 bar bar 1.005 °C 93.045

146.512 °C °C 146.512 93.042 kg/s kg/s 93.042

3.0MW/K 4.409 bar bar 4.409 617.212 kJ/kg kJ/kg 617.212 1.005 bar 1.005 °C 157.864

1.2MW/K

147.162 °C °C 147.162 93.157 kg/s kg/s 93.157

4.824 bar bar 4.824 620.033 kJ/kg kJ/kg 620.033 1.005 bar 1.005 °C 200.642 0.1MW/K 1.006 bar 1.006 °C 202.625 4.824 bar bar 4.824 °C 147.162 182.665 kJ/kg kJ/kg 182.665 kg/s 93.042 161.589 kJ/kg kJ/kg 161.589 kg/s 6069.128 0.9MW/K 1.007 bar bar 1.007 °C 238.721 1.000 bar bar 1.000 °C 30.000 9.902 bar bar 9.902 °C 43.413 0.950 bar bar 0.950 °C 38.557 1.008 bar bar 1.008 °C 247.662 996.307 kJ/kg kJ/kg 996.307 kg/s 0.000 26.1MW/K M 2803.045 kJ/kg kJ/kg 2803.045 kg/s 15.362 1.2MW/K 2425.878 kJ/kg kJ/kg 2425.878 kg/s 96.686 619.976 kJ/kg kJ/kg 619.976 kg/s 93.157 H M kJ/kg kg/s 40.339 bar 40.339 °C 231.254 1.008 bar 1.008 °C 290.931 2603.780 kJ/kg kJ/kg 2603.780 kg/s 3.644 2845.539 kJ/kg kJ/kg 2845.539 kg/s 12.663 0.1MW/K 0.1MW/K 33.557 bar 33.557 °C 240.153 0.088 bar bar 0.088 °C 43.320 4.409 bar bar 4.409 °C 147.155 617.212 kJ/kg kJ/kg 617.212 kg/s 93.042 P T bar °C 1.009 bar 1.009 °C 294.441 156.965 bar 156.965 °C 229.928 152.449 bar bar 152.449 °C 343.457 4.409 bar 4.409 °C 193.696 1.1MW/K 1554.558 kJ/kg kJ/kg 1554.558 kg/s 0.000 4.409 bar 4.409 °C 146.512 C with a 5% of high pressure steam added as Solar Heat 1.010 bar 1.010 °C 350.780 ◦ G 126439.489 kW 118567.682 kW 1.8MW/K 31.854 bar 31.854 °C 285.562 155.372 bar 155.372 °C 335.368 152.449 bar 152.449 °C 343.457 1.010 bar bar 1.010 °C 454.024 0.3MW/K CC 3P-HRSG CC GT 9FA 50Hz type: ºC 45 Temperature: Atmospheric Injection Heat Solar 5% 1.010 bar 1.010 °C 490.248

3052.320 kJ/kg kJ/kg 3052.320 kg/s 68.776 kJ/kg 3045.712 kg/s 84.138 447.858 °C °C 447.858

0.3MW/K 149.785 bar bar 149.785 1.011 bar bar 1.011 °C 545.585 31.854 bar bar 31.854 °C 325.326 bar 4.409 °C 290.095

G 541.858 °C °C 541.858

664.024 kJ/kg kJ/kg 664.024 kg/s 588.395 183627.727 kW bar 30.091 1.012 bar bar 1.012 °C 573.471

1.013 bar bar 1.013 °C 602.330

199929.198 kW

545.467 °C °C 545.467 0.2MW/K 0.3MW/K kg/s 68.776

147.120 bar bar 147.120 3441.255 kJ/kg kJ/kg 3441.255 1455.053 kJ/kg kJ/kg 1455.053 kg/s 588.395 Figure 9.7: Three Pressure CCPP at 45 45.488 kJ/kg kJ/kg 45.488 kg/s 576.988 15.405 bar 15.405 °C 1239.115 1.010 bar 1.010 °C 45.000

9 Standard Three Pressure CCPP with Solar Boosting

345.292 °C °C 345.292 7.289 kg/s kg/s 7.289

155.965 bar bar 155.965 2593.302 kJ/kg kJ/kg 2593.302

78.222 °C °C 78.222 136.915 kg/s kg/s 136.915 43.716 °C °C 43.716 91.915 kg/s kg/s 91.915

9.892 bar bar 9.892 328.245 kJ/kg kJ/kg 328.245 9.892 bar bar 9.892 183.932 kJ/kg kJ/kg 183.932 2410.620 kJ/kg

1.005 bar bar 1.005 °C 94.283

147.799 °C °C 147.799 91.915 kg/s kg/s 91.915

3.0MW/K 4.512 bar bar 4.512 622.755 kJ/kg kJ/kg 622.755 1.005 bar 1.005 °C 158.935

1.2MW/K

148.018 °C °C 148.018 91.952 kg/s kg/s 91.952

4.945 bar bar 4.945 623.725 kJ/kg kJ/kg 623.725 1.005 bar 1.005 °C 202.559 0.1MW/K 1.006 bar 1.006 °C 204.619 4.945 bar bar 4.945 °C 148.018 183.932 kJ/kg kJ/kg 183.932 kg/s 91.915 162.402 kJ/kg kJ/kg 162.402 kg/s 6069.128 0.9MW/K 1.007 bar bar 1.007 °C 240.096 1.000 bar bar 1.000 °C 30.000 9.892 bar bar 9.892 °C 43.716 0.950 bar bar 0.950 °C 38.752 1.008 bar bar 1.008 °C 249.265 1001.407 kJ/kg kJ/kg 1001.407 kg/s 0.000 26.1MW/K M 2802.876 kJ/kg kJ/kg 2802.876 kg/s 15.706 1.2MW/K 2419.973 kJ/kg kJ/kg 2419.973 kg/s 99.204 623.666 kJ/kg kJ/kg 623.666 kg/s 91.952 H M kJ/kg kg/s 41.533 bar bar 41.533 °C 232.338 1.008 bar 1.008 °C 293.353 2593.302 kJ/kg kJ/kg 2593.302 kg/s 7.289 2848.381 kJ/kg kJ/kg 2848.381 kg/s 12.933 0.1MW/K 0.1MW/K 34.432 bar 34.432 °C 241.624 0.089 bar bar 0.089 °C 43.623 4.512 bar bar 4.512 °C 148.010 622.755 kJ/kg kJ/kg 622.755 kg/s 91.915 P T bar °C 1.009 bar 1.009 °C 297.011 160.257 bar 160.257 °C 231.702 155.965 bar bar 155.965 °C 345.292 4.512 bar 4.512 °C 195.290 1.1MW/K 1567.951 kJ/kg kJ/kg 1567.951 kg/s 0.000 4.512 bar 4.512 °C 147.799 C with a 10% of high pressure steam added as Solar Heat 1.010 bar 1.010 °C 352.256 G ◦ 129465.068 kW 121789.103 kW 1.8MW/K 32.652 bar 32.652 °C 287.645 158.752 bar 158.752 °C 337.377 155.965 bar 155.965 °C 345.292 1.010 bar bar 1.010 °C 450.361 0.3MW/K CC 3P-HRSG CC GT 9FA 50Hz type: ºC 45 Temperature: Atmospheric Injection Heat Solar 10% 1.010 bar 1.010 °C 487.217

3041.826 kJ/kg kJ/kg 3041.826 kg/s 70.602 kJ/kg 3039.263 kg/s 86.308 443.084 °C °C 443.084

0.3MW/K 153.157 bar bar 153.157 1.011 bar bar 1.011 °C 542.841 32.652 bar bar 32.652 °C 321.944 bar 4.512 °C 287.096

G 538.796 °C °C 538.796

664.024 kJ/kg kJ/kg 664.024 kg/s 588.395 183627.727 kW bar 30.797 1.012 bar bar 1.012 °C 571.839

1.013 bar bar 1.013 °C 602.330

199929.198 kW

541.684 °C °C 541.684 0.2MW/K 0.3MW/K kg/s 70.602

150.349 bar bar 150.349 3427.449 kJ/kg kJ/kg 3427.449 1455.053 kJ/kg kJ/kg 1455.053 kg/s 588.395 Figure 9.8: Three Pressure CCPP at 45 45.488 kJ/kg kJ/kg 45.488 kg/s 576.988 15.405 bar 15.405 °C 1239.115 1.010 bar 1.010 °C 45.000

9 Standard Three Pressure CCPP with Solar Boosting

346.005 °C °C 346.005 8.747 kg/s kg/s 8.747

157.350 bar bar 157.350 2589.072 kJ/kg kJ/kg 2589.072

79.952 °C °C 79.952 139.384 kg/s kg/s 139.384 43.828 °C °C 43.828 91.384 kg/s kg/s 91.384

9.955 bar bar 9.955 335.503 kJ/kg kJ/kg 335.503 9.955 bar bar 9.955 184.403 kJ/kg kJ/kg 184.403 2405.928 kJ/kg

1.005 bar bar 1.005 °C 95.110

147.834 °C °C 147.834 91.384 kg/s kg/s 91.384

3.0MW/K 4.550 bar bar 4.550 622.909 kJ/kg kJ/kg 622.909 1.005 bar 1.005 °C 159.334

1.2MW/K

148.325 °C °C 148.325 91.470 kg/s kg/s 91.470

4.990 bar bar 4.990 625.050 kJ/kg kJ/kg 625.050 1.005 bar 1.005 °C 203.309 0.1MW/K 1.006 bar 1.006 °C 205.402 4.990 bar bar 4.990 °C 148.325 184.403 kJ/kg kJ/kg 184.403 kg/s 91.384 162.699 kJ/kg kJ/kg 162.699 kg/s 6069.128 0.9MW/K 1.007 bar bar 1.007 °C 240.632 1.000 bar bar 1.000 °C 30.000 9.955 bar bar 9.955 °C 43.828 0.950 bar bar 0.950 °C 38.823 1.008 bar bar 1.008 °C 249.895 1003.382 kJ/kg kJ/kg 1003.382 kg/s 0.000 26.1MW/K M 2802.797 kJ/kg kJ/kg 2802.797 kg/s 15.843 1.2MW/K 2417.704 kJ/kg kJ/kg 2417.704 kg/s 100.131 624.989 kJ/kg kJ/kg 624.989 kg/s 91.470 H M kJ/kg kg/s 42.011 bar bar 42.011 °C 232.757 1.008 bar 1.008 °C 294.311 2589.072 kJ/kg kJ/kg 2589.072 kg/s 8.747 2849.497 kJ/kg kJ/kg 2849.497 kg/s 13.044 0.1MW/K 0.1MW/K 34.781 bar 34.781 °C 242.201 0.090 bar bar 0.090 °C 43.734 4.550 bar bar 4.550 °C 148.317 622.909 kJ/kg kJ/kg 622.909 kg/s 91.384 P T bar °C 1.009 bar 1.009 °C 298.029 161.553 bar 161.553 °C 232.398 157.350 bar bar 157.350 °C 346.005 4.550 bar 4.550 °C 195.907 1.0MW/K 1573.274 kJ/kg kJ/kg 1573.274 kg/s 0.000 4.550 bar 4.550 °C 147.834 C with a 12% of high pressure steam added as Solar Heat 1.010 bar 1.010 °C 352.828 G ◦ 130620.366 kW 123024.062 kW 1.8MW/K 32.969 bar 32.969 °C 288.465 160.082 bar 160.082 °C 338.165 157.350 bar 157.350 °C 346.005 1.010 bar bar 1.010 °C 448.906 0.3MW/K CC 3P-HRSG CC GT 9FA 50Hz type: ºC 45 Temperature: Atmospheric Injection Heat Solar 12% 1.010 bar 1.010 °C 486.018

3037.581 kJ/kg kJ/kg 3037.581 kg/s 71.329 kJ/kg 3036.527 kg/s 87.172 441.220 °C °C 441.220

0.3MW/K 154.484 bar bar 154.484 1.011 bar bar 1.011 °C 541.738 32.969 bar bar 32.969 °C 320.588 bar 4.550 °C 285.817

G 537.567 °C °C 537.567

664.024 kJ/kg kJ/kg 664.024 kg/s 588.395 183627.727 kW bar 31.077 1.012 bar bar 1.012 °C 571.182

1.013 bar bar 1.013 °C 602.330

199929.198 kW

540.148 °C °C 540.148 0.2MW/K 0.3MW/K kg/s 71.329

151.618 bar bar 151.618 3421.844 kJ/kg kJ/kg 3421.844 1455.053 kJ/kg kJ/kg 1455.053 kg/s 588.395 Figure 9.9: Three Pressure CCPP at 45 45.488 kJ/kg kJ/kg 45.488 kg/s 576.988 15.405 bar 15.405 °C 1239.115 1.010 bar 1.010 °C 45.000

65 9 Standard Three Pressure CCPP with Solar Boosting

◦ T( C) Qsolar(MW) psteamturbine (MWe) m˙ HP EV AP (kg/s) m˙ solar (kg/s) 30 8.85 125.27 68.7 3.4 45 8.83 117.98 66.4 3.4 30 17.61 128.02 68.6 6.7 45 17.57 120.65 65.8 6.7 30 21.09 129.03 67.9 8 45 21.04 121.68 65.6 8

Table 20: Three Pressure CCPP with Solar Boosting: solar steam added as CRH.

Percentage Temperature psteamturbine m˙ HP EV AP m˙ solar Qsolar ◦ of nominal steam ﬂow C (MWe) (kg/s) (kg/s) (MW) 5 % 30 125.83 68.6 4 10.59 45 118.5 66.3 4 10.55 10 % 30 129.07 67.9 8.1 21.61 45 121.77 65.6 8.1 21.18 12 % 30 130.05 67.7 9.7 25.53 45 123.05 65.4 9.7 25.46

Table 21: Three Pressure CCPP with Solar Boosting injected as CRH. amount of solar heat the SHP injection improves the steam turbine power output.

For calculating the amount of solar heat needed for heating the condensed water at high pressure until superheated steam at high pressure we use the same procedure than in the case of adding saturated steam. We calculate the enthalpy diﬀerence between both points of measure and multiply it for the mass ﬂow added. The Table 21 shows the results obtained for the second part of the study.

Comparing the Tables 20 and 21 we can conclude that:

• By adding the same amount of solar heat, with the SHP technology the steam turbine delivers more output power than with the CRH technology. That proves that adding steam at high pressure is always better so the eﬃciency of the SHP technology is higher than the CRH eﬃciency.

• By adding the energy needed for increasing the temperature of the feed water to the steam temperature in the point of injection in each case, the SHP technology allows to obtain more output power (130.36 MWe with SHP, while with CRH 130.05 MWe) with the addition of a lower amount of solar heat (21.09 MWth in the case of SHP, while 25.53MWth with the CRH technology). That means that for obtaining the same power output the addition of solar heat in the case of CRH has to be bigger. The consequence of that is the need of more eﬃcient parabolic troughs or a bigger covered collector surface in the solar ﬁeld with the CRH technology.

66 9 Standard Three Pressure CCPP with Solar Boosting

88.981 °C °C 88.981 180.947 kg/s kg/s 180.947 39.378 °C °C 39.378 94.947 kg/s kg/s 94.947

11.723 bar bar 11.723 373.538 kJ/kg kJ/kg 373.538 11.723 bar bar 11.723 165.974 kJ/kg kJ/kg 165.974 2915.829 kJ/kg

1.004 bar bar 1.004 °C 97.264

143.027 °C °C 143.027 94.947 kg/s kg/s 94.947

3.3MW/K 4.761 bar bar 4.761 602.258 kJ/kg kJ/kg 602.258 1.005 bar 1.005 °C 161.272

1.2MW/K

150.011 °C °C 150.011 96.231 kg/s kg/s 96.231

5.228 bar bar 5.228 632.327 kJ/kg kJ/kg 632.327 1.005 bar 1.005 °C 204.829 0.1MW/K 1.006 bar 1.006 °C 206.893 5.228 bar bar 5.228 °C 150.011 165.974 kJ/kg kJ/kg 165.974 kg/s 94.947 143.242 kJ/kg kJ/kg 143.242 kg/s 6069.128 1.0MW/K 1.007 bar bar 1.007 °C 243.944 1.000 bar bar 1.000 °C 25.000 11.723 bar bar 11.723 °C 39.378 0.950 bar bar 0.950 °C 34.167 2802.715 kJ/kg kJ/kg 2802.715 kg/s 9.678 1.008 bar bar 1.008 °C 252.654 1019.260 kJ/kg kJ/kg 1019.260 kg/s 0.000 26.1MW/K M 2802.228 kJ/kg kJ/kg 2802.228 kg/s 15.109 1.3MW/K 35.113 bar bar 35.113 °C 242.748 2387.023 kJ/kg kJ/kg 2387.023 kg/s 104.626 632.263 kJ/kg kJ/kg 632.263 kg/s 96.231 H M kJ/kg kg/s 43.374 bar bar 43.374 °C 236.126 1.008 bar 1.008 °C 293.139 2850.924 kJ/kg kJ/kg 2850.924 kg/s 13.427 0.1MW/K 0.1MW/K 36.761 bar 36.761 °C 245.399 0.071 bar bar 0.071 °C 39.271 4.761 bar bar 4.761 °C 150.003 602.258 kJ/kg kJ/kg 602.258 kg/s 94.947 P T bar °C 1.008 bar 1.008 °C 296.347 155.989 bar 155.989 °C 234.452 4.761 bar 4.761 °C 197.098 1.1MW/K 1551.892 kJ/kg kJ/kg 1551.892 kg/s 0.000 4.761 bar 4.761 °C 143.027 1.009 bar 1.009 °C 350.477 G 137141.092 kW 130045.109 kW 35.113 bar 35.113 °C 287.622 154.268 bar 154.268 °C 334.942 1.8MW/K 1.009 bar 1.009 °C 454.640 151.116 bar 151.116 °C 342.752 C with a 12% of high pressure steam added as Solar Heat with CRH Injection ◦ 0.3MW/K CC 3P-HRSG 50Hz 9FA GT type: ºC 30 Temperature: Atmospheric Injection Reheat Cold 12% 1.010 bar 1.010 °C 492.337

3037.341 kJ/kg kJ/kg 3037.341 kg/s 92.482 449.698 °C °C 449.698

0.3MW/K 148.535 bar bar 148.535 1.011 bar bar 1.011 °C 545.289 4.761 bar bar 4.761 °C 286.491

G 539.585 °C °C 539.585

665.505 kJ/kg kJ/kg 665.505 kg/s 608.469 213307.159 kW bar 32.984 1.012 bar bar 1.012 °C 575.113

1.013 bar bar 1.013 °C 602.313

228545.622 kW

547.309 °C °C 547.309 0.2MW/K 0.3MW/K kg/s 67.695

145.953 bar bar 145.953 3447.485 kJ/kg kJ/kg 3447.485 1478.034 kJ/kg kJ/kg 1478.034 kg/s 608.469 30.317 kJ/kg kJ/kg 30.317 kg/s 596.030 16.009 bar 16.009 °C 1253.936 1.010 bar 1.010 °C 30.000 Figure 9.10: Three Pressure CCPP at 30

67 10 Conversion Eﬃciency of Solar Boosting

10 Conversion Eﬃciency of Solar Boosting

The energy conversion eﬃciency is an useful ratio for comparing the impact of introducing solar heat. We will deﬁne it as the ratio between the additional electric power when we inject the CCPP with solar steam and the amount of solar heat that we are injecting.

Pel,solar − Pel ηsolar = (10.1) Qth,solar We will compare the conversion eﬃciencies for both SHP and CRH technologies at 30◦C.

10.1 Conversion Eﬃciency of Saturated High Pressure Steam For the case of injection with saturated high pressure steam, the conversion eﬃciency is calculated as:

Pel,30◦C,solar − Pel,30◦C 130.36MWe − 122.49MWe ηsolar,30◦C = = = 0.373 (10.2) Qth,30◦C,solar 21.09MWe

The SHP technology has a nominal gross efficiency which is only 1% lower than the efficiencies published in literature for PTC systems (38.5% gross efficiency in [14]).

10.2 Conversion Eﬃciency of Cold Reheated Steam When the injection of solar heat is made in the cold reheat pipe:

Pel,30◦C,solar − Pel,30◦C 130.05MWe − 122.49MWe ηsolar,30◦C = = = 0.296 (10.3) Qth,30◦C,solar 25.53MWe Comparing the results for SHP and CRH we can conclude that the SHP is a more efficient technology. With SHP we obtain 130.36 MWe by injecting only 21.09 MWth of solar heat, while with CRH we need more thermal heat for reaching lower power output. Although the SHP technology is more expensive than CRH in terms of installation in the CCPP, because the injection is done around 155 bar, due to its high efficiency we need a smaller solar field than with the CRH technology being cheaper in terms of construction of the solar field.

11 Impact of Solar Steam Injection on Temperatures and Pressures

11.1 Impact of Saturated High Pressure Steam Injection In this chapter we will discuss brieﬂy the impact of injecting saturated steam at high pressure as solar heat on the temperature of the superheated steam and its pressure. The Table 22 shows the values of pressures and temperatures in each case:

In the Table 22 the pHP,SH,INLET represents the high pressure value of the saturated steam in the point of injection. The THP,ST represents the temperature of the superheated steam before entering the steam turbine.

68 11 Impact of Solar Steam Injection on Temperatures and Pressures

Saturated Steam ◦ Percentage of solar steam added TAMB pHP,SH,INLET (bar) THP,ST ( C) 0% ISO 161 540 0% 153.7 545.7 5% 30 ◦C 157.2 541.9 10% 160.5 538.3 12% 162 536.6 0% 149 549 5% 45 ◦C 152.6 545.3 10% 156 541.7 12% 157.3 540.2

Table 22: Variation of temperature and pressure with the injection of saturated steam at high pressure.

Taking a look at the results we see that the partload operation of the CCPP causes a decrease in pressure. However, the injection of saturated steam at high pressure makes the pressure rise. With the temperatures occurs the opposite, the partload operation leads to a increase in temperatures while the injection of saturated steam results in a decrease. That occurs because when we work in partload at 30◦C for example, the gas turbine is delivering a fixed energy to the HRSG. If we inject new mass flow from the solar the heat exchangers have to heat more mass flow now and the temperature is lower. With the pressure occurs the opposite, with the injection of solar steam the pressure has to increase because of the bigger mass flow in the cycle.

11.2 Impact of Solar Heat Injection as Cold Reheated Steam The Table 23 presents the results for the injection of cold reheated steam at intermediate pressure. THRH represents the temperature in the cold reheat pipe and pRH represents the pressure of reheat, before entering the second stage of the steam turbine.

Saturated Steam ◦ Percentage of solar steam added TAMB pRH (bar) THRH ( C) 0% ISO 32 540 0% 30.4 543 5% 30 ◦C 31.5 541.8 10% 32.5 540.2 12% 33 539.6 0% 29.4 544.8 5% 45 ◦C 30.5 543.5 10% 31.6 542 12% 32 541.4

Table 23: Variation of temperatures and pressures when solar heat is added as cold reheated steam.

69 12 Thermal Eﬃciency of the CCPP

With the injection of cold reheated steam the variation of pressures and temperatures is exactly in the same way than with saturated high pressure steam. But with CRH the variation in both pressure and temperature is smaller compared with SHP as we can see in the Table 24.

Technology ∆P (bar) ∆T (◦C) SHP 8.3 9 CRH 2.6 3.4

Table 24: Variation of temperatures and pressures with solar steam injection

It is worthwhile to mention that the limit of HP value is not attained. The pressure is lower than in the ISO case (155 bar), see Figure 9.10.

12 Thermal Eﬃciency of the CCPP

The thermal eﬃciency of the CCPP is deﬁned as the ratio between the output power obtained from the whole power plant and the energy given to the gas turbine. The energy given to the gas turbine is the amount of fuel burnt in the combustion chamber multiplied by its lower heating value. For the case of an ambient temperature of 30◦C:

PGT + PST 213.3MW + 122.5MW η30◦C = = = 0.541 (12.1) m˙ fuel · Hu 12.4kg/s · 50.015MJ/kg

12.1 Thermal Eﬃciency of the CCPP with Solar Boosting If we assume that the solar energy is for free, we can calculate the thermal eﬃciency of the CCPP when solar heat is injected:

PGT + PST 213.3MW + 130.4MW η30◦C,SHP = = = 0.554 (12.2) m˙ fuel · Hu 12.4kg/s · 50.015MJ/kg

PGT + PST 213.3MW + 130.1MW η30◦C,CRH = = = 0.553 (12.3) m˙ fuel · Hu 12.4kg/s · 50.015MJ/kg The thermal eﬃciency of the CCPP increases with the addition of solar heat and is higher in the case of the SHP technology.

13 Standard Three Pressure CCPP with Increased Solar Boosting

Until now we have studied the introduction of solar energy in a CCPP for improving the overall eﬃciency. We only considered the cases with warmer temperatures and partload performance in the gas turbine.

The steam turbine is designed for given conditions (ISO Conditions) and the design calculations correspond to these conditions. That means that, for example the cross section

70 13 Standard Three Pressure CCPP with Increased Solar Boosting of the steam turbine is fixed in design conditions and it defines the limit of the steam flow through the turbine. If we want to inject more steam in the cycle the total mass never can be bigger than in the design conditions.

Even in ISO conditions (15◦C) it is possible to obtain solar energy with the parabolic trough technology. That means, that at ISO conditions we are wasting the energy obtained in the solar ﬁeld because the limited cross section does not allow to inject more steam in the Rank- ine cycle. Wasting this solar energy makes the solar ﬁeld less economic.

We can consider the injection of solar energy in design conditions by making the cross section of the steam turbine bigger. How big? We have to design a new steam turbine whose mass ﬂow in design conditions is the sum of the solar steam and the steam ﬂow that we had in the previous steam turbine in ISO conditions.

APPROACH ”B”

For modeling this concept in Ebsilon we will use the three pressure cycle with solar boosting. But the diﬀerence now is that we will also introduce steam heated with solar energy in design conditions.

The Table 26 shows the results obtained for the three pressures CCPP when in design conditions we add a 12% of solar heat. We have to notice that the HRSGs in both cycles (with solar or no solar in design conditions) are not exactly the same. In the cycle with solar injection in design conditions, the gas temperature in the economizer at the end of the gas stream is not so low as in the cycle without solar injection. This is because the mass of water through it is bigger and the temperature of the water is higher as well. That makes impossible to reach the temperature of 85◦C at the end of the gas steam. Instead we reach a temperature of 86.7◦C and despite of this small diﬀerence we can consider that both HRSGs work in a similar way.

By comparing the results for ISO conditions, with or without solar heat injection, we can conclude that the addition of solar heat to the CCPP in design conditions allows reaching higher output power. If the hours of operation of the solar ﬁeld are bigger we will obtain more energy for a deﬁned investment costs and the energy will be cheaper in a long-term.

Conditions m˙ HP EV AP (kg/s) m˙ solar (kg/s) psteamturbine (MWe) Qsolar(MW) ISO 72.9 138.5 ISO with Solar Inject. 67.7 8.7 145.96 21.48

Table 25: Comparison between the CCPP working in ISO conditions with and without solar injection.

For all the cases in a range between 10◦C and 45◦C the mass of solar heat injected is the same. When the temperature rises and the gas turbine starts working in oﬀ-design conditions, the production of steam in the high pressure evaporator decreases. The solar heat needed for heating the mass ﬂow decreases with higher temperatures and the result is a lower output power in the steam turbine.

71 14 Economic Analysis of the Three Pressure CCPP with Solar Boosting

12% of solar injection CASES m˙ solar (kg/s) m˙ HP EV AP (kg/s) m˙ total (kg/s) Qsolar(MW) psteamturbine (MWe) design 8.7 67.7 76.5 21.48 145.96 10◦C 8.7 67.8 76.5 21.57 142.90 30◦C 8.7 64.1 72.9 21.32 130.24 45◦C 8.7 61.9 70.7 21.26 122.24

Table 26: Results for a Three Pressure CCPP with 12% of Solar Injection in a range of 10◦C to 45◦C

APPROACH ”C”

Other possibility is the addition of solar heat until reaching the maximum limit of mass ﬂow going through the steam turbine in every case. As we said, when the temperature increases the mass of high pressure steam evaporated is smaller and the mass of solar heat which can be introduced in the cycle until reaching the limit is bigger. The mass ﬂow going through the turbine is in all the cases constant and the solar heat increases with the temperatures. The result is a higher power output at warmer temperatures.

The last conﬁguration makes sense because with warmer temperatures the amount of solar heat available is higher. The reasonable is to add more solar heat at higher temperatures and to improve the overall power output.

CASES m˙ solar (kg/s) m˙ HP EV AP (kg/s) m˙ total (kg/s) Qsolar(MW) psteamturbine (MWe) design 8.7 67.7 76.5 21.48 145.96 10◦C 8.7 67.8 76.5 21.57 142.90 30◦C 16.1 60.4 76.5 38.80 136.69 45◦C 20.6 55.9 76.5 49.35 132.39

Table 27: Results for a Three Pressures CCPP with Maximum Solar Heat Injection in a range of 10◦C to 45◦C

14 Economic Analysis of the Three Pressure CCPP with Solar Boosting

In this chapter a rough economic analysis of the costs for electricity production will be done. We will know if the price of producing electricity in our CCPP with solar boosting is economically competitive between others technologies. The comparison with the expected price and the real price allows us to determine if the project is feasible or not.

For doing the economic study we will assume that the CCPP is already built and currently in operation and due to the development of the solar technologies we want to equip

14 Economic Analysis of the Three Pressure CCPP with Solar Boosting

347.862 °C °C 347.862 8.745 kg/s kg/s 8.745

161.000 bar bar 161.000 2577.625 kJ/kg kJ/kg 2577.625

66.795 °C °C 66.795 145.468 kg/s kg/s 145.468 29.052 °C °C 29.052 100.268 kg/s kg/s 100.268

11.200 bar bar 11.200 280.497 kJ/kg kJ/kg 280.497 11.200 bar bar 11.200 122.804 kJ/kg kJ/kg 122.804 2456.221 kJ/kg

1.003 bar bar 1.003 °C 86.795

149.488 °C °C 149.488 100.268 kg/s kg/s 100.268

3.3MW/K 4.700 bar bar 4.700 630.039 kJ/kg kJ/kg 630.039 1.004 bar 1.004 °C 161.320

1.3MW/K

149.528 °C °C 149.528 100.275 kg/s kg/s 100.275

5.200 bar bar 5.200 630.245 kJ/kg kJ/kg 630.245 1.004 bar bar 1.004 °C 206.147 0.1MW/K 1.005 bar bar 1.005 °C 208.337 5.200 bar bar 5.200 °C 149.528 122.804 kJ/kg kJ/kg 122.804 kg/s 100.268 100.580 kJ/kg kJ/kg 100.580 kg/s 6401.425 0.9MW/K 1.006 bar bar 1.006 °C 242.569 1.000 bar 1.000 °C 15.000 11.200 bar bar 11.200 °C 29.052 0.950 bar bar 0.950 °C 23.962 1.007 bar 1.007 °C 252.186 1010.006 kJ/kg kJ/kg 1010.006 kg/s 0.000 27.5MW/K M 2802.470 kJ/kg kJ/kg 2802.470 kg/s 17.960 1.3MW/K 2323.603 kJ/kg kJ/kg 2323.603 kg/s 109.014 630.177 kJ/kg kJ/kg 630.177 kg/s 100.275 H M kJ/kg kg/s 44.000 bar 44.000 °C 234.160 1.007 bar 1.007 °C 297.965 2577.625 kJ/kg kJ/kg 2577.625 kg/s 8.745 2853.892 kJ/kg kJ/kg 2853.892 kg/s 14.585 0.1MW/K 0.1MW/K 36.000 bar 36.000 °C 244.186 0.040 bar bar 0.040 °C 28.962 4.700 bar bar 4.700 °C 149.519 630.039 kJ/kg kJ/kg 630.039 kg/s 100.268 P T bar °C 1.008 bar 1.008 °C 301.969 166.700 bar bar 166.700 °C 232.305 161.000 bar 161.000 °C 347.862 4.700 bar bar 4.700 °C 198.337 1.0MW/K 1570.624 kJ/kg kJ/kg 1570.624 kg/s 0.000 4.700 bar bar 4.700 °C 149.488 1.009 bar 1.009 °C 355.862 G 152036.621 kW 145955.156 kW 1.8MW/K 34.000 bar 34.000 °C 291.969 164.700 bar 164.700 °C 338.066 161.000 bar 161.000 °C 347.862 1.009 bar 1.009 °C 449.940 CC 3P-HRSG 9FA50Hz type: GT Conditions ISO ofInjection SolarHeat 12% 0.3MW/K 1.010 bar bar 1.010 °C 485.723

3036.153 kJ/kg kJ/kg 3036.153 kg/s 76.474 3039.131 kJ/kg kJ/kg 3039.131 kg/s 94.435 435.420 °C °C 435.420

0.3MW/K 158.000 bar bar 158.000 1.011 bar bar 1.011 °C 538.654 34.000 bar 34.000 °C 321.145 4.700 bar bar 4.700 °C 287.279

G 540.082 °C °C 540.082

667.303 kJ/kg kJ/kg 667.303 kg/s 641.000 259796.323 kW bar 32.000 1.012 bar 1.012 °C 569.382

1.013 bar bar 1.013 °C 602.296

270621.170 kW

540.296 °C °C 540.296 0.3MW/K 0.4MW/K kg/s 76.474

155.000 bar bar 155.000 3418.506 kJ/kg kJ/kg 3418.506 1514.784 kJ/kg kJ/kg 1514.784 kg/s 641.000 15.155 kJ/kg kJ/kg 15.155 kg/s 627.074 17.000 bar 17.000 °C 1278.554 Figure 13.1: Three Pressure CCPP at ISO Conditions with a 12% of high pressure steam added as Solar Heat. 1.010 bar bar 1.010 °C 15.000

14 Economic Analysis of the Three Pressure CCPP with Solar Boosting

347.710 °C °C 347.710 16.064 kg/s kg/s 16.064

160.698 bar bar 160.698 2578.587 kJ/kg kJ/kg 2578.587

80.438 °C °C 80.438 147.174 kg/s kg/s 147.174 39.078 °C °C 39.078 92.174 kg/s kg/s 92.174

10.505 bar bar 10.505 337.583 kJ/kg kJ/kg 337.583 10.505 bar bar 10.505 164.613 kJ/kg kJ/kg 164.613 2415.299 kJ/kg

1.004 bar bar 1.004 °C 94.687

148.824 °C °C 148.824 92.174 kg/s kg/s 92.174

3.3MW/K 4.658 bar bar 4.658 627.178 kJ/kg kJ/kg 627.178 1.005 bar 1.005 °C 160.622

1.3MW/K

149.191 °C °C 149.191 92.238 kg/s kg/s 92.238

5.131 bar bar 5.131 628.787 kJ/kg kJ/kg 628.787 1.005 bar bar 1.005 °C 206.695 0.1MW/K 1.006 bar bar 1.006 °C 208.997 5.131 bar bar 5.131 °C 149.191 164.613 kJ/kg kJ/kg 164.613 kg/s 92.174 142.467 kJ/kg kJ/kg 142.467 kg/s 6401.425 0.8MW/K 1.007 bar bar 1.007 °C 241.651 1.000 bar 1.000 °C 25.000 10.505 bar bar 10.505 °C 39.078 0.950 bar bar 0.950 °C 33.981 1.008 bar 1.008 °C 251.619 1008.588 kJ/kg kJ/kg 1008.588 kg/s 0.000 27.5MW/K M 2802.550 kJ/kg kJ/kg 2802.550 kg/s 17.616 1.3MW/K 2383.435 kJ/kg kJ/kg 2383.435 kg/s 108.238 628.722 kJ/kg kJ/kg 628.722 kg/s 92.238 H M kJ/kg kg/s 43.419 bar 43.419 °C 233.860 1.008 bar 1.008 °C 299.087 2578.587 kJ/kg kJ/kg 2578.587 kg/s 16.064 2855.721 kJ/kg kJ/kg 2855.721 kg/s 14.187 0.1MW/K 0.1MW/K 35.727 bar 35.727 °C 243.746 0.070 bar bar 0.070 °C 38.982 4.658 bar bar 4.658 °C 149.182 627.178 kJ/kg kJ/kg 627.178 kg/s 92.174 P T bar °C 1.008 bar 1.008 °C 303.349 165.260 bar bar 165.260 °C 232.914 160.698 bar 160.698 °C 347.710 4.658 bar bar 4.658 °C 199.091 1.0MW/K 1580.056 kJ/kg kJ/kg 1580.056 kg/s 0.000 4.658 bar bar 4.658 °C 148.824 1.009 bar 1.009 °C 354.748 G 144180.663 kW 136688.190 kW 1.8MW/K 33.803 bar 33.803 °C 293.364 163.668 bar 163.668 °C 339.273 160.698 bar 160.698 °C 347.710 1.009 bar 1.009 °C 442.780 CC 3P-HRSG 50Hz 9FA GT type: ºC 30 Temperature: Atmospheric Injection Solar Limit of C with high pressure steam added as Solar Heat until reaching the mass ﬂow limit in 0.3MW/K ◦ 1.010 bar bar 1.010 °C 479.371

3030.697 kJ/kg kJ/kg 3030.697 kg/s 76.500 3033.212 kJ/kg kJ/kg 3033.212 kg/s 94.116 431.495 °C °C 431.495

0.3MW/K 157.696 bar bar 157.696 1.011 bar bar 1.011 °C 533.816 33.803 bar 33.803 °C 318.766 4.658 bar bar 4.658 °C 284.351

G 536.546 °C °C 536.546

665.505 kJ/kg kJ/kg 665.505 kg/s 608.469 213307.159 kW bar 31.816 1.012 bar 1.012 °C 566.742

1.013 bar bar 1.013 °C 602.313

228545.622 kW

537.897 °C °C 537.897 0.3MW/K 0.4MW/K kg/s 76.500

154.694 bar bar 154.694 3412.212 kJ/kg kJ/kg 3412.212 1478.034 kJ/kg kJ/kg 1478.034 kg/s 608.469 the steam turbine. 30.317 kJ/kg kJ/kg 30.317 kg/s 596.030 16.009 bar 16.009 °C 1253.936 1.010 bar bar 1.010 °C 30.000 Figure 13.2: Three Pressure CCPP at 30

74 14 Economic Analysis of the Three Pressure CCPP with Solar Boosting the present installations with a solar ﬁeld for improving the electricity production with high ambient temperatures and sunny weather.

For designing the collector surface of the solar heat, we will consider two diﬀerent technologies for the solar ﬁeld: Parabolic Trough Concentrators (PTC) and Pneumatic Pre-Stressed Concentrators (PPC).

14.1 Pneumatic Pre-Stressed Concentrators (PPC) The PPC system is currently developed by HELIOVIS AG and Vienna University of Tech- nology and it is still in a prototype phase. The PPC technology consists of two pressurized air chambers separated by a membrane with mirror-coating, [14]. The curvature of the mirror is produced by a higher pressure in the upper air chamber. A picture of the PPC system is shown in the Figure 14.1

The solar ﬁeld will be designed for injecting the CCPP with 12% of mass of steam in

Figure 14.1: Sketch of pneumatic prestressed concentrator, [14] nominal conditions as solar heat. At 30◦C the heat needed for injecting that mass are 21.09 MW of thermal heat (Qth).

Taking into account that the maximum optical eﬃciency of the PPC is ηopt = 0.734 the radiation needed in out solar ﬁeld is:

Qth 21.09 Qrad = = = 28.73MW (14.1) ηopt 0.734 Besides, the speciﬁc power absorbed by the collectors in the solar ﬁeld has a value of 800 W/m2, being the necessary heat absorbed: Q 21.09 · 106 A = rad = = 35916.2m2 (14.2) collector 800W/m2 800 · 0.734 The collector surface needed for reaching the amount of solar heat required are 35916.2 m2. The collector surface is the mirrored surface needed for absorbing the radiation, which is

75 14 Economic Analysis of the Three Pressure CCPP with Solar Boosting different from the total surface of the solar field. The rows of collector have to keep a distance between them because of the shading effects for low incident angles. This distance is approximated as three times the diameter of the collectors and it allows us to determine the land surface necessary for installing the solar field.

Assuming that the land use factor is 3.5, the total surface required for installing the solar ﬁeld is: 2 Asolarfield = Acollector · 3.5 = 125706.7m (14.3) For the study we assume estimated speciﬁc costs for the PPC technology of around 60 e/m2. By multiplying the collector price per square meter and the square meters of collector surface we can calculate the capital expenditure (CAPEX):

CAP EX = 35916, 2m2 · 60e/m2 = 2154972e (14.4)

That is the required initial investment for building the solar field at the beginning of the construction. But for paying the construction during the period of operation of the solar field we have to calculate the annuity. The annuities are the fixed payments which have to be done every year during the predicted life of operation of the solar field for pay off the initial investment. The annuity includes the price of the construction and the increase of interest rate every year.

Considering a period of operation of 15 years (T) and an interest rate of 8% (r), the annuity is calculated as: (1 + r)T · r a = = 0.1168year−1 (14.5) (1 + r)T − 1 And the yearly payment is:

CAP EX · a = 2154972e · 0.1168year−1 = 251700.7e/year (14.6) which in 15 years add makes a total of:

251700, 7e/year · 15years = 3775510.94e (14.7)

The diﬀerence of power that we could obtain by boosting the CCPP with the solar ﬁeld is:

∆P30◦C = Psolar30◦C − P30◦C = 130.36 − 122.49 = 7.87MW (14.8)

∆P45◦C = Psolar45◦C − P45◦C = 123.03 − 115.39 = 7.64MW (14.9) The assumption is that the CCPP is currently in operation and it is situated in Almer´ıa, in the south of Spain. Taking a look at the weather conditions from TRANSYS-Database MEteotest we can know how many days with enough radiation we can expect for running the solar ﬁeld. The minimum incident radiation is assumed to be 600 MW/m2. The number of hours with at least this minimum radiation are shown in Table 28.

By multiplying the number of hours of supposed operation of the solar ﬁeld per year and the energy that we can obtain from it, we know the MWh that the solar ﬁeld delivers per year:

7.87MW · 689hours = 5422.43MW h30◦C (14.10)

76 Almería España - Google Maps http://maps.google.com/maps?f=q&source=s_q&hl=es&geocode=...

14 Economic AnalysisDirección of theAlmería Three Pressure CCPP with Solar Boosting España

◦ ◦ 2 Temperature N hours N hours with Qrad ≥ 600W/m 20 ≤ T < 30 3184 689 T > 30 191 88

Table 28: N ◦hours with enough radiation energy for operating the solar ﬁeld

Figure 14.2: Location of the CCPP, www.Google.com

7.64MW · 88hours = 672.32MW h45◦C (14.11) T otal = 6094.75MW h (14.12) To calculate the cost of electricity (CoE) we assume the operating cost (OPEX) to be equivalent to the 2% of the speciﬁc capital expenditure, CAP EX · a OPEX CoE = + (14.13) P MW h P MW h 1 de 1 11/04/2010 20:08 CAP EX · a 0.02 · CAP EX CoE = + = 48.36e/MW h (14.14) P MW h P MW h Producing one MWh of electricity with the CCPP and the solar ﬁeld using PPC technology cost 4.836 ce/kW h.

14.2 Parabolic Trough Concentrators (PTC) In this chapter the same economic evaluation will be done but now the technology used in the solar ﬁeld are the parabolic trough concentrators (PTC). The CCPP and its location are exactly the same so the calculations for the weather conditions done for PPC are also valid for PTC.

77 15 Economic Analysis of the Three Pressure CCPP with Increased Solar Boosting

The maximum optical eﬃciency for the PTC is ηopt=0.75 and the necessary radiation absorbed by the solar ﬁeld has to be:

Qth 21.09 Qrad = = = 28.12MW (14.15) ηopt 0.5

The speciﬁc power absorbed by the collectors coincides with the case of PPC, 800 W/m2, being the necessary heat absorbed:

Q 21.09 · 106 A = rad = = 35150m2 (14.16) collector 800W/m2 800 · 0.75

And the necessary land extension considering a land use factor of 3.5 are:

2 Asolarfield = Acollector · 3.5 = 123025m (14.17)

For PTC technology the cost of the collector is assumed to be around 280e/m2, [14] and the initial investment required for the construction of the solar ﬁeld:

CAP EX = 35150m2 · 280e/m2 = 9842000e (14.18)

For the same expected period of operation the annual payments are equivalent to:

CAP EX · a = 9842000e · 0.1168year−1 = 1149545.6e/year (14.19) which at the end makes a total of:

251700, 7e/year · 15years = 17243184e (14.20)

And the CoE: CAP EX · a OPEX CoE = + = 220.91e/MW h (14.21) P MW h P MW h The cost of producing one MWh of electricity with the PTC technology is 22.091 ce/kW h, which is more expensive than with conventional technologies or the PPC and for that reason less attractive.

15 Economic Analysis of the Three Pressure CCPP with Increased Solar Boosting

Until now we have done the calculations for knowing the costs of producing one MWh of electricity but we have only considered a range of temperatures between 20◦ C and 45◦ C. That was because we had a built CCPP with a steam turbine and the injection of solar steam was only possible when the mass of the steam turbine was smaller than in nominal conditions. Now, we will consider a built CCPP but our steam turbine is big enough and the injection of solar steam is possible in all the range of temperatures. Even in ISO conditions there are no additional costs of injecting solar steam because all the components of the CCPP can aﬀord an increment of mass ﬂow in the circuit.

The mass of steam injected in the new design case will be 12% of the nominal steam in

78 15 Economic Analysis of the Three Pressure CCPP with Increased Solar Boosting

ISO conditions for the case without solar boosting, which are 8.75 kg/s. The total heat needed for injecting that mass in ISO conditions are 21.48 MW of thermal heat.

The diﬀerence of power that we could obtain in ISO conditions:

∆PISO = PsolarISO − PISO = 145.96 − 138.5 = 7.46MW (15.1) and taking into account that the number of hours with enough radiation for running the solar system are shown in Table 29:

◦ ◦ 2 Temperature N hours N hours with Qrad ≥ 600W/m 0 ≤ T < 20 5386 526 20 ≤ T < 30 3184 689 T > 30 191 88

Table 29: N ◦hours with enough radiation energy for operating the solar ﬁeld

By multiplying the new number of hours of operation of the solar ﬁeld and the power that we could obtain by using it, we know the MW h that the solar ﬁeld produces:

7.87MW · 689hours = 5422.43MW h30◦C (15.2)

7.64MW · 88hours = 672.32MW h45◦C (15.3)

7.46MW · 526hours = 3923.96MW hISO (15.4) T otal = 10018.71MW h (15.5) By taking advantage of the radiation even in ISO conditions we can produce about 3900 MW h more that in the case before by not increasing too much the costs. As we said, the CCPP will be the same and the only thing which increases the investment in the installation is a bigger solar ﬁeld because now it has to absorb more radiation.

15.1 Pneumatic Pre-Stressed Concentrators (PPC) The calculations for the new solar ﬁeld with PPC technology are presented in the following. The values of optical eﬃciencies and land use factors are the same that in the chapter before.

The collector surface needed for absorbing 21.48 MW of thermal heat are:

Q 21.48 · 106 A = rad = = 36580.4m2 (15.6) collector 800W/m2 800 · 0.734 and taking into account the land use factor:

2 Asolarfield = Acollector · 3.5 = 128031.33m (15.7)

The necessary initial investment for performing the construction of the solar ﬁeld:

CAP EX = 36580.4m2 · 60e/m2 = 2194822.89e (15.8)

79 16 Conclusion and Summary

Finally, the cost of the electricity using the bigger solar ﬁeld: CAP EX · a OPEX CoE = + = 29.97e/MW h (15.9) P MW h P MW h

With PPC technology and running the solar field joined the CCPP at all operation temperatures the cost of electricity decreases and it is even more competitive. With this configuration we will obtain benefits from the first day of operation.

15.2 Parabolic Trough Concentrators (PTC) For the PTC technology, the collector surface and the surface needed for building the solar ﬁeld are: Q 21.48 · 106 A = rad = = 35800m2 (15.10) collector 800W/m2 800 · 0.75 2 Asolarfield = Acollector · 3.5 = 125300m (15.11) The initial investmet:

CAP EX = 35800m2 · 280e/m2 = 10024000e (15.12)

And the cost of producing one MW h of electricity: CAP EX · a OPEX CoE = + = 136.87e/MW h (15.13) P MW h P MW h

With PTC technology this cost is higher, because although is a mature and proven technology its speciﬁc cost is still high and compared with PPC technology is not a competitive technology.

16 Conclusion and Summary

As a conclusion of our study, we will discuss the results obtained in the technical and economic analysis for the diﬀerent cases and technologies used. Table 30 shows the comparison between the two proposed technologies.

The main results of the technical analysis were:

• For a 9FA Three Pressure CCPP designed and strictly limited by the ISO case, approximately 7 additional MWe ( 2% of the installed CCPP power) can be obtained at ◦ 30 C, at a gross thermal eﬃciency of 37% by injecting to the steam cycle 21 MWth of solar heat with the saturated high pressure method.

• In the case of the cold reheat steam injected to the cycle, also approximately 7 MWe can be obtained at 3030◦C of ambient temperature, at a gross thermal eﬃciency of almost 30%.

Comparing these results we can conclude that for obtaining the same amount of power output the size of the solar ﬁeld whit the SHP is smaller. We will use the SHP in the construction of the solar ﬁeld project.

80 16 Conclusion and Summary

Cases Power Techn. CAPEX (e) Land surface (m2) CoE(ce/kW h) Solar Boosting 6094.75 PPC 2154972 125706.7 4.836 20◦ C

Table 30: Comparison betwen diﬀerent technologies and conﬁguration for producing electricity

As the PTC technology has a better optical eﬃciency the amount of radiation absorbed per unit of surface is bigger than with PPC technology. That means that the land surface needed for the PTC technology is smaller than for the PPC technology. But the fact that the cost of the PPC technology is much smaller than the PTC technology, makes PPC a more attractive technology due to its higher proﬁtability.

Taking into account that producing one kWh of electricity with conventional technologies costs about 5-6 ce/kW h we can conclude that the PPC allows us to obtain a competitive price of electricity by using a cleaner technology. That means that the construction of the solar ﬁeld for helping the CCPP is a feasible project and the initial investment will be re- covered plus beneﬁts.

Besides, running the solar field for a range of temperatures between 0◦ C to 45◦ C, allows us to obtain higher radiation with approximately the same costs. The difference between solar boosting and the increased solar boosting is the more expensive solar field due to its wider extension, but with the benefits of producing more solar energy and reducing the costs of electricity production.

81 References

References

[1] Saravanamuttoo,H.I.H., Rogers, G.F.C., Cohen H., Gas Turbine Theory, Pearson Edu- cation, Cornwall, 2001.

[2] Fundamentals of Gas Turbines, William B., John Wiley and Sons, Inc., United States of America, 1996.

[3] Boyce, M., Handbook for Cogeneration and Combined Cycle Power Plants, Asme Press, New York, 2002.

[4] Boyce, M., Ph.D and P.E., Gas Turbine Engineering Handbook, Butterworth- Heinemann, Oxford, 2006.

[5] Ganapathy, V., Industrial Boilers and Heat Recovery Steam Generators, Marcel Dekker,Inc., New York, 2003.

[6] Petchers, N., Combined Heating, Cooling and Power Handbook, Marcel Dekker, Inc., New York, 2003.

[7] Lechner, C., Seume, J., Station¨are Gasturbinen, Springer, Berlin, 2003.

[8] Wengenmayr,R., Buhrke,¨ T., Renewable Energy, Wiley-VCH, Weinheaim, 2008 . [9] Abengoa Solar, www.abengoasolar.com, 2010.

[10] Birnbaum, J., Eck M., Fichtner, M., Hirsch, T., Lehman, D., Zimmermann, G., A Direct Steam Generation Solar Power Plant with Integrated Thermal Storage, 2008.

[11] Herrmann, U., Nava, P., Performance of the Skal-et Collectors of the Andasol Power Plants

[12] Zarza, E., López, C., Cámara, A., Mart´ınez, A., Burgaleta, J. I., Mart´ın, J. C., Fresneda, A., Almer´ıa GDV The first Solar Power Plant with Direct Steam Generation.

[13] General Electric Energy, www.gepower.com, 2010

[14] Hartl, M., Ponweiser, K., Haider, M., Tiefenbacher, F., Comparison of a Prestressed Pneumatic Concentrator CSP System with Conventional Systems.

[15] Willinger, R., Thermische Turbomaschinen, Lecture notes, TU Wien, 2008.

[16] Dolezal, R., Kombinierte Gas und Dampfkraftwerte, Springer Verlag, Berlin, 2001

[17] Steiner, A., Numerische Prozesssimulation von Thermischen Energieanlagen, Lecture Notes, 2009.

[18] Elsaket G., Simulating the Integrated Solar Combined Cycle for Power Plants Applica- tion in Libia, Cranﬁeld University, 2007.

[19] Desertec-UK, www.trec-uk.org.uk, 2010

[20] Kreuzer D., Dynamic Simulation of a Combined Cycle Power Plant, ITE Department of Thermodynamics and Thermal Engineering, TU Vienna, 2008.

82 References

[21] Eﬀenberger, H., Dampferzeugung, Springer Verlag, Berlin, Heidelberg, New York, 2000.

[22] Wikipatents, www.wikipatents.com, 2010.