HHO driven CCPP

by
Gabriele Galli
HHO energy
Introduction:

The objective of the thermodynamics HHO thermal power plants is the determination and
maximization of the efficiency of combined cycle power plants. With other words, this work applies
the methods for increasing the plant effectiveness by saving costs fuel.
In the production of electricity, vapour, and heatingcooling 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 project we propose a CCPP 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 PWVM particulated
water and vapour mixture retrofit can deliver additional steam for the steam cycle without spending the
investment costs of power plant. Therefore the concept is 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. 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 presented and compared: or
with cold reheated steam (CRH) Combined recycle heating. The chosen solution will be analyzed both
technically and economically comparing two concentrating thermal technologies, With the available
technologies only for started all the PPCC.
1Combined 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 more 66%. In design conditions, the gas turbine supplies the 60% with only a
power generator, while the steam turbine delivers only the 34% of the energy.
The combined cycle exists in different engine 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 HHO gasses in some form of
useful power, such as mechanical power. “adiabatic process “.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 HHO gasses and air whit water particulated at high pressure are mixed, and a turbine where
the mixture of HHO gases are expanded after the combustion. In the turbine, the gases are expanded
“adiabatically” with 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,
relationship:
and depends only on the pressure ratio (r) and the nature of the gas (γ is the adiabatic 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 temperature , we can
obtain the following equation:

(2.2)
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]
The 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:
(2.3)
dQ = dH - V dp - p dV + p dV
dQ = dH - V dp
and thermodinamics is a branch of physics that deals with the energy and work of a system. In aerodynamics,
we are most interested in thermodynamics in the study of propulsion sistems and understanding high flows.
The first law of thermodynamics indicates that the total energy of a system is conserved. Total energy includes
the potential and kinetic energy, the work done by the system, and the transfer of heat through the system. The
secon law of thermodynamics indicates that, while many physical processes that satisfy the first law are
possible, the only processes that occur in nature are those for which the entropy of the system either remains
constant or increases.

Entropy, like temperature and pressure can be explained on both a macro scale and a micro scale. Since
thermodynamics deals only with the macro scale, the change in entropy delta S is defined here to be the heat
transfer delta Q into the system divided by the temperature T:
delta S = delta Q / T
During a thermodynamic process, the temperature T of an object changes as heat Q is applied or extracted. A
more correct definition of the entropy S is the differential form that accounts for this variation.
dS = dQ / T
The change in entropy is then the inverse of the temperature integrated over the change in heat transfer. For
gases, there are two possible ways to evaluate the change in entropy. We begin by using the first law of
thermodynamics:
dE = dQ - dW
where E is the internal energy and W is the work done by the system. Substituting for the definition of work for a
gas.
dQ = dE + p dV
where p is the pressure and V is thevolume of the gas. If we use the definition of the enthalphy H of a gas:
H = E + p * V
Then:
dH = dE + p dV + V dp
Substitute into the first law equation:
dQ = dH - V dp - p dV + p dV
dQ = dH - V dp
is an alternate way to present the first law of thermodynamics. For an ideal gas, the equation of state is written:
p * V = R * T
where R is the gas constant. The heat transfer of a gas is equal to the heat capacity times the change in
temperature; in differential form:
dQ = C * dT
If we have a constant volume process, the formulation of the first law gives:
dE = dQ = C (constant volume) * dT
Similarly, for a constant pressure process, the formulation of the first law gives:
dH = dQ = C (constant pressure) * dT

If we assume that the heat capacity is constant with temperature, we can use these two equations to define the
change in enthalpy and internal energy. If we substitute the value for p from the equation of state, and the
definition of dE in the first energy equation, we obtain:
dQ = C (constant volume) * dT + R * T dV / V
Similarly substituting the value of V from the equation of state, and the definition of dH we obtain the alternate
form:
dQ = C (constant pressure) * dT - R * T dp / p
Substituting these forms for dQ into the differential form of the entropy equation gives::
dS = C (constant volume) * dT / T + R * dV / V
and
dS = C (constant pressure) * dT / T - R * dp / p
These equations can be integrated from condition "1" to condition "2" to give:
S2 - S1 = Cv * ln ( T2 / T1) + R * ln ( V2 / V1)
and
S2 - S1 = Cp * ln ( T2 / T1) - R * ln ( p2 / p1)
where Cv is the heat capacity at constant volume, Cp is the heat capacity at constant pressure, and ln is the
symbol for the logarictim function.
If we divide both equations by the mass of gas, we can obtain intrinsic, or "specific" forms of both equations:
s2 - s1 = cv * ln ( T2 / T1) + R * ln ( v2 / v1)
and
s2 - s1 = cp * ln ( T2 / T1) - R * ln ( p2 / p1)
where cp and cv are the specific heat capacity depending on the type of process we encounter, we can now
determine the change in entropy for a gas.
These equations can be a bit confusing, because we use the specific heat at constant volume when we have a
process that changes volume, and the specific heat at constant pressure when the process changes
pressure.To clarify matters, let's look at the first equation:
s2 - s1 = cv * ln ( T2 / T1) + R * ln ( v2 / v1)
If we have a constant volume process, the second term in the equation is equal to zero, since v2/v1 = 1. We can
then determine the value of the specific heat for the constant volume process. But if we have a process that
changes volume, the second term in the equation is not zero. We can think of the first term of the equation as
the contribution for a constant volume process, and the second term as the additional change produced by the
change in volume. A similar type of argument can be made for the equation used for a change in pressure.

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
Figure 2.3: Pressure Ratio for Maximum Work per kg of Air.

Regeneration
The regeneration effect takes place by adding a heat exchanger which heats the air at its entrance
reducing the HHO consumption. The efficiency can be increased because the water- 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.
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
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 effciency of the cycle.
Is possible to simulate the Carnot cycle. The efficiency cycle is given by: CARNOT
(2.4) and when the speciﬁc heats are constant, we can assume this configuration:
(410.3)
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 (the efficiency).
will decrease because in that part the effciencies of the Carnot cycles are smaller.
An intercooling regenerative cycle can increase the power output and the thermal effciency. This
combination provides an increase in effciency of 12% and an increase in power output of about 30%,
Pascal law:
In cases of perfect gases H2 and O2 the follow state is :
Reheting

at
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 diagram.

There is an intermediate pressure which divides the stages for both process, reheat and intercooling.
2.1.2 Part-load performance
The variation of thermal efficiency with reduction in power, sometimes referred to as part-load
performance, is of major importance in applications where considerable running at low power settings
is required,
When determining the 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 .

decrease in both power output and efficiency.
The start point for studying the ﬂow rate behaviour of a multistage turbine is the investigation 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:(2.7)

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, (2.8)
Combining equations 2.9, we obtain:

(2.10)
For the sound pressure ratio

(2.12)
(2.13)
If the pressure ratio decreases below the critical value compare to the equation (2.12), the
7
capacity does not change, that means that the capacity function ξ remains at the.

By leaving κ = constant, the relative capacity function:
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,

(2.15)
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.
(2.16)
2 Combined Cycle Power Plants:

Figure 2.8: Relative flow rate function, [15]
Taking into account the caloric equation of state pAva = κ−1 κ hA, we get:
(2.17)
For a multi-stage turbine the influence of the revolution speed is small, so we can state ¯ µ/ ¯ µ0 = 1.
Finally, we write the simplified equation as:
(2.18)

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, with ˙ m 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.
(2.19)

In the turbine condenser the final pressure has only a hundredth bar, that means that pratically 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. The performance of the steam turbine is described by the Rankine cycle.
9
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:
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 formation. 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.
where ˙ ms 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.
Figure 2.11: Regenerative and reheated Rankine cycle,

where ˙ mair and ˙ mw are respectively the mass of air and water and the the subscripts ”o”and 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):
(2:23)
where 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 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 .
Air-cooled Condenser

12
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.
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.

Of the heat is transferred:
The majority 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.

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.
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: the efficiency of the
cycle. There are some vertical HRSGs with natural circulation or the other possibility 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.
One two tree
Water is being delivered to the drum with the help of the pumps. In the drum, the
14
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

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,
(2.24)
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.
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.
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 diffeerence 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 diagram for one section of the HRSG where we can see the evolution of
the water and steam going through the HRSG:
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 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.
To different temperature.
Diagram above: GT Performance at different ambient temperatures. Given for LM 5000 gas turbine.

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.

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
Figure 2.15: HRSG performance versus ambient temperature.
Gas ﬂow shown has a multi- plication factor of 0.1, [5]

Pessure
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 a power increase.

Energy Distribution in a Combined Cycle Power Plant,
tent at the exit of the steam turbine. The increase in moisture content creates les erosion and corrosion
problems in the later stages of the turbine. A limit is set at about 10% moisture content, Boyce
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.
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,
Fig: IAPWS-IF h-p-t

FIG,t:p:h:s
nominal 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:
• Hose objects
Components and pipes form cycles, while value crosses, text fields, graphical elements andHose
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
4 Modeling with EBSILONr Professional
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
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.
We will use the Design and Partload models in studying the impact caused for feeding the plant with
energy coming from a solar field.

5.1 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
?
The 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.
G
GE Turbine 9FA 50HzISO Conditions 17.000 1278.0771514.160 641.000 P bar
T °C H kJ/kg M kg/s
17.002 420.710 436.334 627.07417.002
15.000 32.681 13.926 1.013 15.00015.155 627.074
1.013 602.00 666.952 641.000
260017.690 kW
270851.761 kW
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 effciencies 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
?
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 Of-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 for each case. Knowing the density values obtained in each case are showed in
the
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.
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.
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:
Figure 5.4: Variation of gas turbine performance in T-s diagram depending on the ambient temperature,
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
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.
60%
65%
70%
75%
80%
85%
90%
95%
100%
105%
110%
110%
100%
90%
80%
70%
60%
50%
40%
30%
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.
Figure 5.6: Comparison between the T,s diagrams for the gas turbine accepting less air.
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 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 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
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.
GE Turbine 9FA 50Hz
ISO Conditions
80% of Nominal Power
15.097 1228.503 1445.996 578.568 P bar T °C H kJ/kg
M kg/s 15.099

T.403.385 417.655 566.580
15.099 15.000
32.681 11.988
1.013 15.000
15.155 566.580
1.013 602.007
665.534 578.56
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 Effciency 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 effciency of the gas
turbine.
Table 4: Variation of Parameters in Off-design when the power is speciﬁed
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.
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 varieties 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

Table 8: Variation of Heat Transfer Capacity with the mass flow on the steam/water pipe.
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
6 Modeling a Single Pressure CCPP
?
Table 11: Stodola Correction for the Steam Turbine
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 chap ters ”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 de-aerator 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 de-aerator 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. The super heater 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 super heaters 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 super heater 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.
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 super heaters.
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 de-aerator. The de-aerator is the component which guarantees that there is
no air and other dissolved gases going from the feed water to the HRSG.. The water which enters the
de-aerator 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 de-aerator 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 HHO.
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 Temperature is 10◦C.. 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 effciency of the
generator are equivalent to 112.8 MWe.
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

Figure6.1:CCPPwithaSinglePressureHRSG
Profile: Design
01 Secondary (Heat_exchanger) 02 Secondary (Heat_exchanger_4) 03 Secondary (Heat_exchanger_1)
04 Secondary (Evaporator) 05 Secondary (Heat_exchanger_2)
06 Secondary (Heat_exchanger_3) 07 Secondary (Heat_exchanger_5) 08 Primary (Heat_exchanger)
09 Primary (Heat_exchanger_4) 10 Primary (Heat_exchanger_1)
11 Primary (Evaporator) 12 Primary (Heat_exchanger_2) 13 Primary (Heat_exchanger_3) 14 Primary
(Heat_exchanger_5)
Temperature (°C)
0
100
200
300
400

500
600
700
P ower (kW)
0 50000
1 100000
2 150000
3 200000
4 250000
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 part load for generators and pumps is
Table 13: Variation of Electrical Efficiency in Generators in Part load
7 Two Pressure CCPP

Table 14: Variation of Isentropic Efficiency in Pumps in Part load
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 super heater at low pressure, two re heaters 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 de-aerator, 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 super heaters 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.
CCPPTwo Pressures Gas Turbine Power (MW e) 260 Steam Turbine Power (MW e) 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
HkJ/kg Mkg/s
260017.690 kW
124385.152 kW
270851.761 kW
129567.867 kW
Figure7.1:CCPPTwoPressureHRSG
8 Modeling a Three Pressure CCPP
q,T-Diagram 2P_DEFINITIVE.ebs
Profile: Design
01 Secondary (Heat_exchanger) 02 Secondary (Heat_exchanger_5) 03 Secondary (Heat_exchanger_4)
04 Secondary (Heat_exchanger_1) 05 Secondary (Evaporator) 06 Secondary (Heat_exchanger_2) 07
Secondary (Heat_exchanger_6) 08 Secondary (Evaporator_1) 09 Secondary (Heat_exchanger_7) 10
Secondary (Heat_exchanger_3) 11 Secondary (Heat_exchanger_8) 12 Primary ( Heat_exchanger ) 13
Primary (Heat_exchanger_5) 14 Primary (Heat_exchanger_4) 15 Primary (Heat_exchanger_1)
Temperature (°C) Power (kW)
0 0
100 50’000
200 100’000
300 150’000
400 200’000
500 250’000
600 300’000

Figure 7.2: Heat transfer in the HRSG with Two Pressure CCPP
The steam after the low pressure evaporator is sent to the de-aerator.
In the three pressure model, the water after leaving the de-aerator is divided into three pipes. One part
of the water (about the 15%) is evaporated at low pressure and comes back to the de-aerator. 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
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.
Looking at the results we can conclude that the addition of steam coming from the nebulized water and
steam always increases the steam cycle power output because more steam is being expanded in the
steam turbine.
By adding the same percentage of steam heated by nebulized water and steam, the output power is
bigger at much lower temperatures. That is because the HHO gases 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 more steam nebulized water .
9 Standard Pressure CCPP with Steam and nebulized water.
Mount of HHO gases with Steam and nebulized water in the 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 of HHO gases with Steam and particulated water technology the steam

turbine delivers more output power without pollution . That proves that adding steam and water 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 separated the combustible of the feed water to the steam temperature
with water particulated in the point of injection in each case, the SHP technology allows to obtain more
output power
(130.36 MW e with SHP, with CRH 130.05 MW e)
(21.09 MW in the case of SHP, while 25.53MWth with the CRH technology).
11 Impact of water and steam Injection on Temperatures and Pressures
Saturated Steam Percentage of steam added
(◦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 steam of saturated steam at high
pressure.
Taking a look at the results we see that the part load operation of the CCPP causes a decrease in
pressure. However, the injection steam of saturated steam at high pressure makes the pressure rise.
With the temperatures occurs the opposite, the part load operation leads to a increase in temperatures
while the injection of saturated steam results in a decrease. That occurs because when we work in part
load at 30◦C for example, the gas turbine is delivering a fixed energy to the HRSG. If we inject new
mass flow generated you 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.
12 Thermal Efficiency of the CCPP
With the injection of cold reheated steam and water nebulisation the variation of pressures and
temperatures is exactly in the same way than with saturated high pressure steam.
Table 24: Variation of temperatures and pressures with cold steal water nebulisation, injections:
It is worthwhile to mention that the limit of HP/Kw value is not attained. The pressure is lower than in
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:
η30◦C =
PGT + PST ˙ HHO gases ·= Hu
=
213.3MW + 122.5MW 12.4kg/s · 50.015MJ/kg
= 0.541
(12.1)
PGT + PST .HHO · Hu=213.3MW + 130.1MW 12.4kg/s · 50.015MJ/kg= 0.553 (12.3)

The thermal efficiency of the CCPP increases with the addition of . WPCW water particulated cold
vapour and is higher in the case of the SHP technology.
13 Standard Three Pressure CCPP with Increased Until now we have studied the introduction. WPCW
water particulated cold vapour of energy in a CCPP for improving the overall efficiency. We only
considered the cases with warmer temperatures and partload performance in the water gas
turbine. WPCW water particulated/cold vapour.
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.
13 Standard Three Pressure CCPP with Increased water particulate WPCW cold vapour.
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.
That means, that at ISO conditions we are wasting the energy obtained in the electromagnetic field
because the limited cross section does not allow to inject more steam in the Rankine cycle. Wasting this
electromagnetic energy makes the solar field less economic origin from the same systems.
We can consider the injection of HHO energy in design conditions by making the cross section of the
steam turbine bigger. We have to design a new steam turbine whose mass flow in design conditions is
the sum of the an. WPCW water particulated/cold vapour 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 HHO energy in design conditions.
de. WPCW water particulated cold vapour design conditions, the HHO gases temperature in the
economizer at the end of the gas stream is not so low as in the cycle without 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 WPCW water particulated/cold vapour.
heat injection, we can conclude that the addition h. WPCW water particulated/cold vapour.
Eat to the CCPP in design conditions allows reaching higher output power. If the hours of operation of .
WPCW water particolated/ cold vapour are bigger we will obtain more energy for a defined investment
costs and the energy will be cheaper in a medio-term.
The needed for heating the mass flow decreases with higher temperatures and the result is a lower
output power in the steam turbine.
14 Economic Analysis of the Three Pressure CCPP with WPCW water particolated/ cold vapour.
12% of vapour and particulated water.
CASES HHO(kg/s) ˙
HPEV AP (kg/s) ˙
total (kg/s)
(MW) power steam turbine (MW . WPCW water particolated/ cold vapour.

0◦C
8.7
67.7
76.5
21.48
145.96
10◦. WPCW water particolated/ cold vapour.
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 of WPCW Injection in a range of 10◦C to
45◦C
APPROACH ”C”
Other possibility is the addition of WPCW water particulated cold vapour 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 WPCW
water particulated cold vapour 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 WPCW water particulated
cold vapour. Heat vapour reduce with the temperatures. The result is a higher power output with
reduced temperatures.
The last configuration makes sense because with colder temperatures the amount available is WPCW
water particulated water particulated The reasonable addictions is to add less heat at a WPCW water
particulated cold vapour to improve the overall power output.
CASES WPCW (kg/s) ˙ mHPEV AP (kg/s) ˙ mtotal (kg/s) (MW) power steam turbine WPCW Table 27:
Results for a Three Pressures CCPP with Maximum HHO Injection in a range of 10◦C to 45◦C
14 Economic Analysis of the Three Pressure CCPP with of WPCW
In this chapter a rough economic analysis of the costs for electricity production will be done. We know
if the price of producing electricity in our CCPP is economically competitive between other
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 WPCW water particulated cold vapour

Conclusion and Summary
As a conclusion of our presentation, we will present you the results obtained in the technical and
economic analysis for the different cases and technologies used. The comparison between the two
proposed technologies.
The main results of the technical analysis were:
• For a 9FA CCPP designed and strictly limited, 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 is WPCW
water particolated cold to the steam cycle.
• In the case of the cold reheat steam injected to the cycle, also approximately 7 MWe can be obtained
at 30◦C of ambient temperature, at a gross thermal efficiency of almost 30%.
Cases Power Tech
Comparison between different technologies and configuration for producing electricity
As the is WPCW water particulated cold vapour 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 less
5-6 ce/kWh we can conclude that the WPCW water particulated cold vapour allows us to obtain a
competitive price of electricity by using a cleaner technology. That means that WPCW water
particulated/ cold vapour the CCPPis a feasible project and the initial investment will be recovered
plus benefits.

References
[1] Sara vanamuttoo, H.I.H., Rogers, G.F.C., Cohen H., Gas Turbine Theory, Pearson Education,
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., Stationäre Gasturbinen, Springer, Berlin, 2003.
[8] Wengenmayr,R., Bührke, 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., Ca´mara, 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 CSPSystem 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.
[21] Effenberger, H., Dampferzeugung, Springer Verlag, Berlin, Heidelberg, New York, 2000.
[22] Wikipatents, www.wikipatents.com, 2010.

HHO driven CCPP

More Related Content

What's hot (20)

Viewers also liked (10)

Similar to HHO driven CCPP (20)

HHO driven CCPP