13

that in tropical and subtropical regions 15% power output enhancement may be

achieved (Wen and Narula, 2000). Ondryas et al. (1991) presented that the cost of

installing a gas turbine or combined cycle plant rated at 35 °C is 20-30% higher than

that rated at 6.6 °C, so cooling the compressor inlet to a certain temperature will reduce

the installing cost. In their assessment they had also emphasized compressor inlet

temperature limits which should not be below 0 °C to prevent ice building on the

compressor blades since the chilled air shall be at 100% relative humidity due to the

moisture condensation during the air cooling process.

The performance of combined gas turbine plant under three different blade cooling

schemes: air cooling, open circuit steam cooling (OCSC) and closed loop steam cooling

(CLSC) was discussed by (Najjar et al., 2004), and it was reported that more power is

created when the cooling steam in the closed loop is not thrown away, thus, the power

output with (CLSC) is increased by 6%, accompanied by 19% rise in the efficiency

relative to (OCSC) at similar conditions. On the other hand, (CLSC) results in 11%

enhancement in power and 3.2% in the efficiency relative to air cooling at the same

circumstances.

The efficiency of 60% became achievable by Westinghouse and others through an

improvement in operating parameters for both steam and gas turbine and raising

Turbine Inlet Temperature (TIT) to 1427 °C (Bannister et al., 1995, a and b). A

reduction in TIT by 10%, drops the combined specific power by 24% and the efficiency

by 7.1% with air cooling scheme, but if CLSC is used, the decrease in the former will

be 22% and 5.1% in the latter at 10% reduction in TIT (Najjar et al., 2004).

Cogeneration utilizing waste heat recovery is also a well-established technology and

has been implemented in many engineering applications. For example, cogeneration

system, which supplies all heating, cooling and part of the electric power for Great
 14

Western’s computer and office complex in Northridge, Calif., was, put into operation.

This system consists of three gas-turbine generator sets rated at 560 kW each with one

being used as stand-by. One unit operates continuously to supply the cooling needs of

the computer center. Waste heat from the turbine exhaust at 482 oC is used to power a

355-ton double-effect absorption (it has a COP 40% higher than the single-effect

design) chiller in-addition to providing domestic hot water and space heating. The

second unit is operated 8-10 hours a day during work days to supply cooling to the

office building through another 355 ton absorption chiller.

A life cycle cost analysis showed the cogeneration system having a payback period

of 5.1 years with a 19.25 percent rate of return, which equates to a dollar savings of

more than 250,000 $/year (Freeman, 1983). To obtain maximum power output during

hot weather, the inlet air to each turbine is cooled to 15 oC, considering that the most

recently installed power generation unit of a single gas turbine of 70 MW capacity (Al-

Bortmany, 2002, a).

The total capital cost of the 316.8 MW Pasadena cogeneration plant without inlet

cooling (or turbine inlet cooling) is estimated to be $ 237.6 million on ISO conditions

basis (Punwani et al., 2001). Furthermore, the combined plant power output decreases

from 317 MW to 273 MW when the ambient temperature increases to 35 °C.

Therefore, at 35 °C, the same investment of $ 237.6 million increases the effective

capital cost per MW of the uncooled cogeneration power plant from $ 750,000 to about

$870,000. Cooling turbine blades in cogeneration plants improve the overall efficiency

by 2% which may reflect a saving of $ 30-40 million in fuel costs over a typical 30-

year life of a 400-500 MW combined power plant (Najjar et al., 2004).

It was found by (Zurigat et al., 2004) that the average Net Present Value (NPV) of

installing fogging systems at the gas turbine inlet, in two different locations in Oman,
 15

is $ 4,622,018 in Marmul and $ 6,182,496 in Fahud. Also the purchase and installation

costs are paid pack within the first six months of system operation and for later years,

annual revenue of over $ 500,000 is as well generated.

Based on the literature review presented in the foregoing (see Table 3.1) it is seen

that most studies have considered a single or utmost two inlet air cooling techniques

together and very seldom coupled with economic assessments. Furthermore, very few

studies have considered hour-by-hour simulations a matter very important in realistic

assessments of GT power plant performance. Therefore, this work is designed to fill the

gap in the literature by considering a number of GT inlet air cooling techniques and

conduct life cycle cost analysis under local weather conditions of one location in

Jordan.

Table 3.1: Literature review summary

Gas Turbine Engine

Author Boosting Technique Result
Gasparovic and
Evaporative after-cooling 55% increase in the power output
Stapersma (1973)
A 10.16 cm H2O pressure drop results in a
Ondryas et al. 1.45% decrease in power output, 0.45%
Chilled water coolers
(1991) increase in heat rate, and an increase in
exhaust gas temperature of 1.1 °C
Ondryas et al. Absorption and mechanical 30% higher output power at -6.7 °C than at
(1991) chillers 35 °C
1.8-3.0 % increase in power output
Guinn (1993) Evaporative cooling
depending on climatic conditions
Guinn (1993) Chilled water coolers 5.7-7.7 % increase the power output
A reduction of inlet air temperature by 28
°C increases power output by 30% and
Kolp et al. (1995) Technique is not specified
decreases heat rate by 4.5% for a 40 MW
General Electric LM6000 GT
De Lucia et al.
Direct evaporative cooling 2% increase in gas-turbine efficiency
(1996)
Najjar and
Evaporative after-cooling 13% increase in the efficiency
Zaamout (1996)
Bartolini and Salvi 8% increase in power output and 4%
Absorption chillers
(1997) increase in the efficiency
Inlet air-cooling from 38 to 4 oC resulted in
Stewart (1998) Technique is not specified 12 MW (21%) of net power increase where
the cooling system uses ice storage
 16

Table 3.1: Literature review summary (cont.)

Gas Turbine Engine

Author Boosting Technique Result
Providing an inlet temperature of 6 oC and
Stewart (1998) Technique is not specified increasing the capacity from the ISO rating
of 30 MW to 37.5 MW (a 25% increase)
At design conditions of 36 oC dry bulb and
24 oC wet bulb temperatures a 3.5 MW
Stewart (1998) Evaporative cooling
increase in capacity of 48 MW gas turbine
at ISO conditions is achieved
Reducing the inlet air temperature to 7 oC
enhanced power output by 20% and 14 %
Aqua-ammonia absorption for the two gas-turbine types used. These
Bortmany (2001)
refrigeration values were found to increase during the
summer months where power augmentation
of 39% and 33% were reported.
1% increase in power output is achieved for
Elliot (2001) Water chillers
every 1.6 °C drop in inlet air temperature
Anon. (2002 and Directing a controlled spray of
10-20 % increase in the power output
2001) water into the compressor inlet
Bassily (2002 and
Indirect evaporative cooling 3% increase in gas-turbine efficiency
2001)
Anon. (2002, a) Direct fogging 10% increase in the power output
Anon. (2002, b) Direct fogging 20% increase in the power output
Mercer (2002) Evaporative cooling 10-15 % increase in the power output
Reduction in inlet air temperature by 3-15
Al-Hazmy and
Evaporative coolers °C, enhances power output by 1-7% and
Najjar (2004)
improves the efficiency by 3%
10% increase in power output during cold
Al-Hazmy and
Vapor compression humid conditions and by 18% during hot
Najjar (2004)
humid conditions
13%increase in power output while
Ameri et al. (2004) Fogging cooling
efficiency improvement was less than 1%
20% increase in power output based on the
Zurigat et al.
Evaporative cooling month with the highest number of
(2004)
evaporative cooling degree hours
Power output was improved for a gas
Zurigat et al. turbine of 40 MW by 20% based on the
Fogging cooling
(2004) month with the highest number of
evaporative cooling degree hours
Combined Power Cycle
Cost of installing a gas turbine or combined
Ondryas et al. Absorption and mechanical
cycle plant rated at 35 °C is 20-30% higher
(1991) chillers
than that rated at 6.6 °C
Bannister et al. Raising Turbine Inlet
The efficiency of 60% became achievable
(1995, a and b) Temperature (TIT)
Wen and Narula 15% increases in power output may
Technique is not specified
(2000) achieved in tropical and subtropical regions
6% increases in power output with closed
loop steam cooling, accompanied by 19%
Najjar et al. (2004) Blade cooling
rise in the efficiency relative to open circuit
steam cooling at similar conditions
11% increase in power output and 3.2% in
Najjar et al. (2004) Blade cooling efficiency with closed loop steam cooling
relative to air cooling at same circumstances
 17

CHAPTER 4

MATHEMATICAL MODELS DEVELOPMENT

In this Chapter the mathematical models of the cycles described in Table 1.1 are

developed. The governing equations are presented with reference to the empirical or

theoretical foundations of the parameters involved. The simple gas turbine model is

described first (see Section 4.1) followed by the simple steam turbine model (see

Section 4.2). Then combined cycle model is developed (see Section 4.3) followed by

the cogeneration cycle model (see Section 4.4).

4.1. Simple Gas Turbine Model with Humidity Effect

Simple gas turbine cycle consists of three main components; air compressor,

combustion chamber (C-C) and gas turbine (see Fig. 4.1). Ambient air is compressed in

the compressor in an adiabatic process 1-2 (or 1-2S if it were isentropic process). In the

combustor fuel is injected and burned under isobaric process 2-3, and an increase in

temperature of the gas mixture is achieved at point 3 (see Fig. 4.2). A slight pressure

drop takes place in the combustion chamber, so the pressure at point 3 is smaller than

the pressure at point 2. Finally, the flow is expanded in the turbine, in an adiabatic

process 3-4 (or 3-4S if it were isentropic process), where part of the energy is extracted

to drive the compressor, and the reset is utilized to produce a rotational motion that

could be used in generating electricity or shaft power. The pressure at turbine exit

(point 4) is greater than atmospheric pressure. Also, the turbine exhaust temperature is

very high so the flue gases can rise to a high level in the atmosphere, hence, a reduction
 18

in the environmental and health impacts is achievable. This cycle is termed the gas

turbine simple cycle. Mathematical models of different gas turbine components are in

order:

2 3
Combustor

Electrical
Generator

Compressor Gas turbine

4
1

Ambient Exhaust gas

Air

Fig. 4.1. Simple gas turbine cycle.

P3
Temperature (K)

P2
2S 2

4
P4 4S
P1
1

Entropy (kJ/kg.K)

Fig. 4.2. Temperature-entropy diagram of simple gas turbine.

 19

4.1.1. The air compressor

The specific work of the compressor (kJ/kg) can be calculated using the first

law of thermodynamics as follows:

WC = C P , a (T02 − T01 ) + d 01 ( hVap , 02 − h Vap , 01 ) (4.1)

The compressor power (kW) is given by (Hameed, 1996):

N C = m& air WC (4.2)

In the above two equations C P , a is the specific heat of air cross the compressor

at constant pressure (kJ/kg.K), d 01 is the humidity ratio at the compressor inlet

(kgwater/kgdry, air), Τ01 is the compressor inlet temperature (in K), Τ02 is the

compressor outlet temperature (in K), m& air is the mass flow rate of the air

(kg/s), hVap , 02 and hVap , 01 are the enthalpies (kJ/kg) of superheated water vapor

at air compressor outlet and inlet, respectively.

The isentropic efficiency of the compressor is given in terms of the

compression ratio RC as (Korakianitis and Wilson, 1994):

⎛ RC − 1 ⎞
η C = 1 − ⎜ 0.04 + ⎟ (4.3)
⎝ 150 ⎠

The compressor outlet temperature (in K) assuming an isentropic efficiency is

given in terms of RC , Τ01 and the specific heat ratio γ air (Hameed, 1996) as:

T01 ⎛ γ air −1 ⎞
T02 = T01 + ⎜ R γ air − 1⎟ (4.4)
ηC ⎝ C ⎠

The air specific heat at constant pressure C P , a (kJ/kg.K) is given in terms of

the average inlet-outlet temperature (in K) (Al-Hazmy and Najjar, 2004) as:
 20

C P , a = 1.0189134 E + 03 − 1.3783636 E − 01T +

(4.5)
+ 1.9843397 E − 04 T 2 + 4.2399242 E − 07 T 3

This is valid within the range 200-800 K. For 800-2200 K range the

following relation is recommended:

C P , a = 7.9865509 E + 02 + 5.3392159 E − 01T −

(4.6)
− 2.2881694 E − 04 T 2 + 3.7420857 E − 08 T 3

The humidity ratio (specific humidity) is given in two forms by Helal (1991)

as

m& Vap
d= (4.7)
m& air

0.622 PVap
d = (4.8)
P − PVap

Where PVap and P are the vapor partial pressure in the moist air and the total

pressure, respectively (in kPa). The vapor partial pressure is calculated by

Sharabati (1994) as:

PVap = RH PSat .Vap (4.9)

Where RH is the relative humidity and PSat .Vap is the saturated vapor partial

pressure (kPa) given by Jones (1994) as:

PSat .Vap = 10 Z (4.10)

Where Z is given as a function of temperature T (in K) by the following relation:

3142.31
Z = 30.59051 − 8.2 log10 T + 0.0024804 T − (4.11)
T

The total pressure at air compressor intake is P01 = Pamb − Δ PAir , Duct

(Hameed, 1996). Where Pamb , ΔPAir , Duct are ambient pressure and pressure drop

across the intake duct (kPa). In this work pressure drop was ignored in simple
 21

gas turbine calculations and had been taken into account for cooled gas turbine

as the inlet air cooler poses an appreciable pressure drop. The mass flow rate

of dry air m& air (kg/s) is given in terms of the volumetric flow rate of dry air

Vol air (m3/s) and the specific volume of dry air ν air (m3/kgdry, air) is given by

Lucas (2003) as:

Volair
m& air = (4.12)
ν air

Tamb
ν air = (0.287 + 0.462 d amb ) (4.13)
Pamb

In the previous equation, the ambient temperature Tamb is in Kelvin and the

ambient pressure Pamb is in kPa. d amb is the ambient air humidity ratio and can

be calculated using Eq. (4.8).

4.1.2. The combustion chamber

Specific released heat from combustor (kJ/kg) is given by Hameed (1996):

q = (1 + d + f ) C P , g (T03 − T02 ) (4.14)

Also, the released heat rate (kW) from the combustion chamber will be:

Q = m& air q (4.15)

In the above equations T03 is the turbine inlet temperature (in K), C P , g is the

specific heat of flue gases across the combustor at constant pressure (kJ/kg.K),

given by Al-Hazmy and Najjar (2004) in terms of the average temperature T

(in K):

C P , g = 1.0887572 E + 03 − 1.4158834 E − 01T +

+ 1.9160159 E − 03T 2 − 1.2400934 E − 06T 3 + (4.16)
+ 3.0669459 E − 10T − 2.6117109 E − 014T
4 5
 22

The fuel air ratio f (kgfuel/kgdry, air), defined by Hameed (1996) as:

m& f
f = (4.17)
m& air

Where m& f is the mass flow rate of the fuel (kg/s).

Energy balance equation can be written for an insulated chamber, and the fuel

air ratio is determined as follows:

C P , g (T03 − 298) + d (hVap , 03 − hVap , 02 ) − C P , a (T02 − 298)

f = (4.18)
Δhc − C P , g (T03 − 298)

Where hVap , 03 is superheated water vapor enthalpy at the combustor outlet and

Δhc is the fuel calorific value, both in (kJ/kg). The constant 298 in Eq. (4.18)

is a reference temperature (25 °C).

Regarding equation (4.14), it should be noted that the flue gases (the

mixture of air, water vapor and fuel) behave as an ideal gas in such cases, so

the partial pressure of each component is approximately equal to the total

pressure. Therefore, all the results are concluded by using total pressure. Also,

there is not humidification or dehumidification process in the power cycle.

Thus, humidity ratio d is constant at the compressor, combustor and the

turbine. The combustor inlet total pressure P02 (kPa) is given by (Molqy,

2000):

P02 = RC P01 (4.19)

Due to the pressure drop in the combustor, the outlet total pressure of

combustion chamber P03 (kPa) is given by Hameed (1996) in terms of the

combustion chamber efficiency η Combustor as:

P03 = P02 η Combustor (4.20)

 23

The total mass flow rate (kg/s) cross the combustor and the turbine can be calculated

based on mass balance as follows:

m& total = m& air (1 + d + f ) (4.21)

4.1.3. The gas turbine

Gas turbine specific work (kJ/kg) is given by Molqy (2000) as:

WT = (1 + d + f ) C P , T (T03 − T04 ) (4.22)

Also, the turbine power (kW) is given as:

N T = m& air WT (4.23)

Where C P , T is the flue gases specific heat (kJ/kg.K), calculated using equation

(4.16) at the turbine inlet-outlet average temperature, T04 is gas turbine outlet

temperature (K).

The isentropic efficiency of the gas turbine is estimated by Korakianitis and

Wilson (1994) as:

⎛ RC − 1 ⎞
η T = 1 − ⎜ 0.03 + ⎟ (4.24)
⎝ 180 ⎠

Gas turbine outlet temperature T04 (K) is determined by Cohen (2004) as:

γ gas −1
⎡ ⎤
⎢⎛ 1 ⎞ γ gas
⎥
T04 = T03 + η T T03 ⎢⎜⎜ ⎟⎟ − 1⎥ (4.25)
R
⎢⎣⎝ C ⎠ ⎥⎦

Where γ gas is the specific heat ratio of the flue gases across the turbine. The

outlet total pressure P04 (kPa) from the turbine is described by Cohen (2004)

as:

1
⎛T ⎞ η T ⎛⎜ γ gas −1 ⎞⎟
P04 = P03 ⎜⎜ 04 ⎟⎟ ⎜⎝ γ gas ⎟⎠ (4.26)
⎝ T 03 ⎠
 24

As stated earlier the gas turbine exhaust pressure and temperature are above

those of the ambient surrounding. A minimum exhaust temperature limit of 90

°C was assumed to minimize the environmental impacts (MEC s 15th WWW,

2004).

The specific net power (kJ/kg) of the gas turbine plant can be evaluated as:

SPGT = WT − WC (4.27)

The net power (kW) of the gas turbine installation is given as:

PGT = N T − N C (4.28)

Gas turbine power cycle efficiency is estimated by Al-Hazmy and Najjar

(2004) as follows:

SPGT
η GT = (4.29)
f Δhc

Or by the following common relation:

PGT
η GT = (4.30)
Q + N Aux

It should be noted that η GT calculated using the above expressions has

different behaviors. Using Eq. (4.29), η GT decreases if the ambient

temperature increases irrespective of the relative humidity. In contrast, using

Eq. (4.30) will result in an increase in efficiency if the ambient temperature

increases keeping the relative humidity equal zero while it may increase or

decrease depending on the relative humidity value. In this research Eq. (4.29)

is used to determine the efficiency as it is believed to be more accurate. The

specific fuel consumption SFC (kgfuel/kWh) is given by Hameed (1996) as:

3600 m& f
SFC = (4.31)
PGT
 25

4.2. Simple Steam Turbine Power Cycle Model

Steam turbine power plants had been used widely in the past. Steam plant is

characterized by high efficiency ranging 40-55%. In spite of its high efficiency, steam

turbine power stations have an inherent disadvantage, the production of high pressure

and temperature involves bulky and expensive steam generating equipments (Cohen,

2004). Simple steam turbine power cycle consists of three main components, steam

generator or boiler, steam turbine and condenser (see Fig. 4.3). As shown in Fig. 4.4 the

water is pumped in an adiabatic or isentropic process (5-6 or 5-6S, respectively). In the

boiler heat is added under isobaric process (6-7 or 6S-7), water vapour is formed at

elevated temperature and pressure (point 7). Pressure drop can take place in the boiler

furnace and in pipes, so the pressure at point 7 will be at 7`. Finally, the steam is

expanded in the steam turbine in an adiabatic or isentropic process (processes 7-8 or 7-

8S, respectively). Expansion process can be 7`-8 or 7`-8S if the pressure drop is taken in

consideration. The condenser works under vacuum constant pressure 8-5 or 8S-5. This

cycle is called simple steam turbine cycle or Rankin cycle if all the processes

mentioned above were reversible (Hameed, 1993). The mathematical models

describing different steam turbine components are in order.

4.2.1. The steam turbine and the condenser

The specific work of a steam turbine (kJ/kg) is given by Hameed (1993) as:

W ST = ( h07 − h08, S ) (4.32)

Also, the steam turbine power (kW) is evaluated by:

N ST = m& steamWST (4.33)

Where m& steam is mass flow rate of steam via steam turbine (kg/s), h07 and h08, S

are steam turbine inlet and outlet enthalpies in an isentropic process (kJ/kg).
 26

Steam
Turbine
Electrical
7´ Generator

Hot Side
8
7
Steam
Condenser

Boiler

Cooling Tower
Water
Pump

Cold Side
6 5

Water Water Basin

Pump
Valve
Makeup water Blow down water

Fig. 4.3. Simple steam turbine power plant.

P7
P7`
Temperature (K)

7`
6
6S
P8

5 8S 8

Entropy (kJ/ kg.K)

Fig. 4.4. Temperature-entropy diagram of simple steam turbine power station.

 27

In ideal case, h07 = f( P07 , T07 ), h08, S = f( P08 , S 08, S ). Also, P08 = 3.0 to

′ = f( P07′ , T07′ ), where the pressure

85.0 kPa (Molqy, 2000). In actual case, h07

drop ΔPBoiler , Turbine from the boiler inlet to the steam turbine inlet is

recommended to be ΔPBoiler , Turbine = ( 2 − 3)% P06 (Molqy, 2000). Where P06 is

the pressure at the boiler inlet and P07′ = P06 − ΔPBoiler , Turbine is inlet pressure of

steam turbine. All thermodynamic properties were calculated using a

subroutine designed to replicate the thermodynamic steam tables (Dawoud,

2003).

The steam turbine outlet enthalpy h08 (kJ/kg) is calculated by the following

relation, given by Helal (1991) as:

′ − η ST (h07
h08 = h07 ′ − h08, S ) (4.34)

Where η ST is steam turbine efficiency which assumed to be within the range

η ST = (80 - 90)% (Hameed, 1993).

From first law of thermodynamics, the condenser load (kW) in ideal case is

estimated by:

QST , Cond = m& steam ( h08, S − h05 ) (4.35)

Where h05 is the enthalpy (kJ/kg) at the condenser exit or at the water pump

entrance, which is determined by h05 = f ( P08 , χ 05 ). χ 05 is the dryness fraction

at saturated wet vapor line (or at the saturation liquid line).

In the actual case, the condenser load is evaluated as:

Q ST , Cond . = m& steam ( h08 − h05 ) (4.36)

Cooling tower mathematical model will be explained in details in Section 4.3.7.

 28

4.2.2. The boiler and water pump

Boiler load (kW) in ideal case can be determined by Molqy (1996):

Q Boiler = m& steam ( h07 − h06 , S ) (4.37)

It can be written for actual cycle as:

′ − h06 )
QBoiler = m& steam (h07 (4.38)

Boiler inlet enthalpy in an isentropic process h06 , S (kJ/kg) is calculated using a

subroutine developed by h06 , S = ( P06 , S 06, S ) , where S 06, S is the outlet water

pump entropy in an isentropic process. On the other hand, h06 the boiler inlet

enthalpy (kJ/kg) in an adiabatic process is given by Hameed (1993) as:

h06, S − h05
h06 = + h05 (4.39)
η w, pump

Where η w, pump is water pump efficiency; η w, pump = (70 − 88)% (Rostom, 1997).

The water pump specific power (kJ/kg) in ideal cycle case is given by Molqy

(2000):

W w, pump = h06, S − h05 (4.40)

Similarly, in the actual cycle case it will be:

Ww, pump = h06 − h05 (4.41)

Ww, pump may also be expressed as (Molqy, 2000):

ν 05 ( P06 − P05 )
Ww, pump = (4.42)
η w, pump

Since the inlet water pump pressure P05 is equal to the condenser pressure, the

water vapor specific volume (m3/kg) at the water pump is ν 05 = f ( P05 , χ 05 ).

 29

Water pump power (kW) can be written as:

N w, pump = m& steam Ww, pump (4.43)

Steam turbine plant net power (kW) is calculated as:

PST = N ST − N w, pump (4.44)

The efficiency of the steam turbine power plant is given as:

PST
η STP = (4.45)
QBoiler

Simple steam turbine power station design can be verified by the following

ratios which are given by Molqy (2000). The first ratio is the condenser to the

boiler load ratio which is described to be within the range 45-50% of that if the

pressure drop and the irreversible losses were relatively low, or it may exceed

50% for tough circumstances. The other ratio is the water pump specific work

to steam turbine specific work. It is given to be within the range 3-4%.

4.3. Combined Cycle Model with Inlet Air Cooling Using Steam
Powered Absorption Refrigeration

Combined cycle (gas and steam turbines) is a technique which is extensively used to

enhance overall efficiency and power output of power plants. Utilizing the waste

exhaust gas from gas turbine installation to heat up the boiler or heat recovery steam

generator (HRSG) will greatly decrease the fuel consumption required for the steam

power plant, hence the power plant efficiency will increase. Steam turbine power

stations have efficiency of about 40-55%, gas turbine 30-34% and cogeneration

technique 55-63% (Hameed, 1996). It should be noted that in this study actual

combined cycle is modeled, i.e., pressure drop and irreversibility are taken into

account.

The combined cycle consists of gas turbine power station and steam power plant

operating as follows: the exhaust gases with moderate pressure P04 and high

temperature T04 supply heat to the HRSG, i.e., economizer, steam generator and

superheater (see Fig. 4.5). Water enters the pump at state 10, and it is compressed

isentropically along the process 10-11S or adiabatically along 10-11 to the HRSG inlet

pressure P11 (see Fig. 4.6). Water enters the economizer with the parameters T11 , P11

and leaves it at T5 , P5 where the water state at the steam generator inlet is considered to

be wet saturated vapor. Change in phase takes place in the steam generator and dry

saturated vapor is obtained at the steam generator exit with T6 , P6 . Then steam (or dry

saturated vapor) is driven to the superheater resulting in outlet temperature and pressure

T7 , P7 . The superheated steam with the parameters T7 , P7 enters the steam turbine

where it expands isentropically along 7-8S process or adiabatically through 7-8 process.

Shaft work is produced which drives an electric generator.

 The pressure and the temperature drop during the expansion process to state 8S-8

where the steam enters the condenser (see Fig. 4.6). The steam at the condenser inlet is

usually considered to be with a high quality (high dryness fraction) and isobaric process

takes place in the condenser 8-9. Condensation process is obtained by rejecting heat to

cooling tower or any other cooling medium. Finally steam leaves the condenser with

low quality and enters the pump (process 9-10) thereby, closing the cycle.

In the considered technique mentioned above, inlet air cooling system is used to cool

the air compressor intake and so, the overall efficiency is boosted. Using an absorption

chiller or any other inlet air cooling technique will decrease the specific work of the

compressor since the cooled air has higher density than uncooled air. Although a

parasitic power is generally needed to drive the cooling machines, this will slightly

affect overall efficiency (Hameed, 1996). In this work the absorption chiller is powered

by steam bled from the steam turbine, i.e., small part of steam is extracted from the

steam turbine with the values T12 , P12 to power the absorption chiller generator. Inlet

and outlet conditions of the bled steam have the values T15 , P15 and T16 , P16

respectively will maintain the generator temperature to be constant.

The steam exiting the generator is fed into a heat recovery heat exchanger of cross-

flow mixed stream type (state T17 , P17 ) where it is mixed with that coming from the

condenser (state 9) where it exits at state 10. This will obviously decrease the specific

work of the pump and also the fuel needed for the HRSG, so an increase in the power

output and the efficiency are achievable.

The absorption chiller consists of generator, condenser, evaporator, solution pump,

and two throttling (expansion valves) as shown in Fig. 4.5. In the generator the solution

(LiBr-H2O or NH3-H2O) is warmed up in an isobaric process to the generator

temperature and pressure ( TGEN , PGEN ). The strong solution (high concentration of H2O
in the former and high concentration of NH3 in the latter) will leave the generator in

vapor state and passes through the condenser while the weak solution with ( TGEN , PGEN

) will be expanded via weak solution throttling valve, then it will enter the absorber in

low temperature and pressure ( T ABSOR. , PABSOR. ). The strong solution enters the

condenser in vapor state and leaves it in liquid state under condensation temperature

and pressure ( TCOND , PCOND ). The high pressurized strong solution in liquid state is

expanded in throttling or an expansion valve to low temperature and pressure ( TEVAP ,

PEVAP ) and it remains in liquid state. It should be noted that the generator and the

condenser operate under same pressure ( PGEN = PCOND ) while the evaporator and the

absorber operate under lower pressure PEVAP which is equal to PABSOR. . The strong

solution enters the evaporator (constant pressure) in liquid state and leaves it in a vapor

state to enter the absorber. A mixture of vapor strong solution and liquid weak solution

(the former comes from the evaporator and the latter from the generator) enters the

absorber where a heat rejection process takes place resulting in both solutions being in

liquid phase. Finally, the solution pump will elevate the pressure of the mixture

(solution) to the generator pressure PGEN .

 13 11 Electrical
7 Generator

Steam Generator Economizer

Steam Turbine
Heat Recovery Steam Generator
12
8
5 Hot Side
19

Hot Side
Steam Bleeding
Steam
Condenser
6
Super heater

18

Cold Side
Water
14 Pump 17
10 9
Water Cooling
Exhaust Cross flow H.E Pump Tower
Gas Mixed Stream

Standard Basin free

Electrical Elbow
4
2 C-C 3 Generator Basin

Blow down valve

Gas Turbine
1 Compressor
Ambient Make-up valve

Expansion
Valve Condenser 16 15
7´ 6´ 5´
Weak Solution

4´
Expansion
Evaporator

1´ Valve
Solution
Pump
2´
3´
8´
Ambient Air

Absorbe

Strong Solution
Ambient

Fig. 4.5. Combined power station with inlet air cooling using absorption refrigeration
powered by steam bled from steam turbine.
 14

Terminal
temperature
18
difference

7
P (bar)
19

Pinch
point
13
5

6
Temperature (K)

11

11S

10

9 8S 8

Entropy (kJ/kg.K)

Fig. 4.6. Temperature-entropy diagram of combined power station with inlet air
cooling using absorption chiller powered by steam bled from steam turbine.
4.3.1. Compressor Inlet Air Cooling Load Calculations

Gas turbine inlet air cooling load (kW), both sensible and latent, can be

estimated by Sharabati (1994):

QC , load = m& air {C P , a , H . E (Tamb − TComp , in ) + h fg , H . E (d amb − d Comp , in )} (4.46)

Where C P , a , H . E is the specific heat of the air cross the heat exchanger (H.E) at

constant pressure (kJ/kg.K) which is given by Eqs. (4.5) and (4.6), TComp , in is

the cooled air temperature at compressor inlet (K), h fg , H . E is the latent heat of

evaporation (kJ/kg) evaluated at the average inlet-outlet temperature across the

cooler (evaporator hot side). d amb d Comp , in are ambient and cooled inlet air

humidity ratios (kgwater/kgdry, air).

The latent heat of evaporation of water h fg , H . E is calculated based on

thermodynamic relations (Dawoud, 2003) which is estimated at the mean

temperature of ambient and inlet air compressor temperatures. Humidity ratio

is estimated using Eqs. (4.7) or (4.8).

The state of cooled air at the compressor inlet is considered to be fully

saturated with constant temperature (Dawoud et al., 2004). The latent load is

taken in consideration if ( d amb − d Comp , in ) is more than zero, otherwise the

cooling load will be sensible.

4.3.2. Cooled Gas Turbine Mathematical Model

The same mathematical model used in simple gas turbine with humidity

effect in Section 4.1 applies here with a slight difference. Pressure drop at the

air duct intake is assumed constant and is taken into account with respect to

the kind of cooling unit used to cool the compressor inlet air. For example, 64
Pa for evaporative coolers, 10 Pa for fogging coolers, 80 Pa for Aqua

Ammonia absorption chillers, 80 Pa for Lithium Bromide-Water absorption

chillers and 80 Pa for electrical coolers (Lucas, 2003). Also, relative humidity

and temperature at compressor inlet assumed to be constant in this research. It

should be noted that the GT compressor inlet temperature must be above 5 °C

to avoid ice forming at the compressor entrance and avoid the structural

damages (Dawoud et al., 2004). Compressor inlet temperature had been

proposed for both Aqua Ammonia and Lithium Bromide-water chillers, to be 8

°C, pressure drop 80 Pa and relative humidity 100%.

Specific net power, total net power and efficiency augmentation of gas

turbine had been determined as follows:

SPC , GT − SPGT
SPAugm = (4.47)
SPGT

PC , GT − PGT
PAugm = (4.48)
PGT

η C , GT − η GT
η Augm = (4.49)
η GT

4.3.3. Steam Turbine and Condenser Mathematical Model:

Steam turbine net power is given in cogeneration cycle by Hameed (1996)

to be PST = (30 − 44)% PGT . This may serve as an initial guess or starting

point. Steam turbine specific work (kJ/kg) with steam bleeding is evaluated by

the first law of thermodynamics as:

WST = (h07 − h012 ) + (1 − χ bled )(h012 − h08 ) (4.50)

 Superheater exit temperature T07 is assumed according to the electrical load

and the pressure is calculated as P07 = P06 − ΔPSH , where pressure drop in the

superheater ΔPSH is recommended to be within the range ΔPSH = (5 − 10)% P06

(Molqy, 2000). Steam generator exit pressure P06 is described by

P06 = P05 − ΔPSG , where pressure drop in steam generator ΔPSG is given within

the range ΔPSG = (2 − 6)% P05 (Molqy, 2000). The economizer exit pressure

P05 is recommended to be P05 = P011 − ΔPECON , where pressure drop in the

economizer ΔPECON is given by Najjar et al. (2004) to be within the range used

in the steam generator, i.e., ΔPSG = (2 − 6)% P05 . Economizer inlet pressure

P011 is assumed to be known as well as the condenser pressure P08 which is

given as P08 = 3.0 to 85.0 kPa (Hameed, 1993).

Condensation temperature T08 is obviously known to be the saturated

temperature at the corresponding pressure, so the enthalpy at the condenser

inlet h08 is given by Molqy (1996) as:

h08 = h07 − η ST ( h07 − h08, S ) (4.51)

Where η ST is the steam turbine efficiency, recommended by Hameed (1993)

to be η ST = 80 - 90 % . Steam turbine outlet enthalpy in an isentropic expansion

h08, S is evaluated by Helal (1991) as:

′′ + (1 − χ 08, S ) h08
h08, S = χ 08, S h08 ′ (4.52)

Where ′ , h08
h08 ′′ and χ 08, S are wet saturated vapor, dry saturated vapor

enthalpies and dryness fraction in an isentropic expansion respectively

evaluated at the condenser pressure.

 The extracted steam ratio from the turbine χ bled and its enthalpy h012 are

assumed to be defined. The enthalpy of the bled steam h012 is chosen with

known temperature and pressure; h012 = f (T012 , P012 ) that can cover the

absorption chiller generator load and estimated as follows:

h012 = h07 − η ST ( h07 − h012 , S ) (4.53)

The mass flow rate of water vapor (kg/s) passing through the plant is given by

Molqy (2000) as follows:

PST
m& steam = (4.54)
[( h07 − h012 ) + (1 − χ bled ) (h012 − h08 )]η mech η elec

Where η mech is the mechanical efficiency of steam turbine which takes into

account the friction losses in bearings, lubricant material and controlling

systems. The mechanical efficiency is assumed to be within the range

η mech = 90 to 99 % and the electrical efficiency of the electric generator η elec ,

which takes into account the electrical losses, is assumed to be equal to η mech

(Molqy, 2000).

The condenser outlet (point 9) state (see Fig. 4.5 and 4.6) is defined by

dryness fraction ′ , both

χ 09 = 0 and wet saturated vapor enthalpy h09 = h09

related to condensation pressure or temperature. The condenser load (kW) can

be written as:

QST , Cond = m& steam (1 − χ bled )( h08 − h09 ) (4.55)

The combined cycle efficiency is given by Molqy (2000) as follows:

If (QRe q − Qav ) > 0 then

PGT + PST
η combined = (4.56)
Q + QRe q − Qav + N Aux

Otherwise
 PGT + PST
η combined = (4.57)
Q + N Aux

Where Qav is the available amount of heat (kW) and N Aux the power

consumed in the auxiliaries (kW).

4.3.4. Cross Flow Heat Exchanger-Mixed Stream and Water Pump

Mathematical Model
The extracted steam leaving the absorption chiller generator with state 16,

still has energy that can be recovered by the water (point 10), i.e., at water

pump intake (see Fig. 4.6). A cross flow heat exchanger-mixed stream was

suggested by Hameed (1993) and Molqy (2000). Two streams were considered

as inputs and one as an output (see Fig. 4.5). The inlet streams are condensed

steam coming from the condenser (state 9) with mass flow rate (kg/s):

m& 09 = m& steam (1 − χ bled ) (4.58)

The second stream is the extracted steam from the steam turbine after

heating the absorption chiller generator (state 17) shown in Fig. 4.5. The

enthalpy h017 = f (T017 , P017 ) and the mass flow rate (kg/s) is:

m& 017 = χ bled m& steam (4.59)

The outlet stream is defined by h010 = hmix and assumed pressure P010 = Pmix

which must be less than pressure at point 17. So, the suction process can take

place in cross flow heat exchange-mixed stream. Steam at states 17 and 9 are

absorbed in the exchanger. The enthalpy at the pump inlet h010 is determined

by Helal (1990):

m& 017 h017 + m& 09 h09

h010 = (4.60)
m& 017 + m& 09
Inlet and outlet entropies of water pump in an isentropic compression is known

to be equal, defined by S 010 = S 011, S = f ( P010 , h010 ) . The pressure at water

pump discharge is assumed equal to the economizer inlet pressure P011 , so the

enthalpy at discharge in an isentropic compression is determined by h011, S = f

( P011 , S 011, S ) . The enthalpy at water pump outlet or at the economizer inlet

h011 can also be evaluated by the following equation (Najjar et al., 2004):

h011 = h010 +
1
η w, pump
(h
011, S − h010 ) (4.61)

Where water pump efficiency is described as η w, pump = 70 to 88% (Rostom,

1997). Water pump specific work (kJ/kg) is given by Najjar et al. (2004) as:

ν 010 ( P011 − P010 )

Ww, pump = (4.62)
η w, pump

Water pump power (kW) can be written as:

N w, pump = m& steam Ww, pump (4.63)

Molqy (2000) had recommended the ratio of water pump specific work to

steam turbine specific work to range from 3 to 4%. By this relation, the design

of steam turbine and water pump specific work can be validated. It was

emphasized by Molqy (2000) that the temperature difference between water

pump outlet and inlet should not exceed the limit of 3 °C. It will be seen in the

next Section that simultaneous validation for steam turbine power plant design

can be done via another equation, which is a function of condenser and HRSG

(or boiler) loads.

4.3.5. Cold Side Heat Recovery Steam Generator Mathematical Model

Steam generator inlet and outlet temperatures T05 , T06 respectively are

assumed to be saturated so the enthalpies corresponding to points 5 and 6 are

estimated, since the pressure and temperature are defined. The recovery load

(kW) on the cold side or the required heat in steam turbine power installation

is given by Molqy (1996) as:

QRe q = Q ECON + QSG + QSH (4.64)

Where Q ECON is the economizer load, QSG is the steam generator load, and

QSH is the superheater load. These loads (kW) are calculated from the first law

of thermodynamics as:

QSH = m& steam (h07 − h06 ) (4.65)

QSG = m& steam (h06 − h05 ) (4.66)

QECON = m& steam (h05 − h011 ) (4.67)

Where h05 is the enthalpy at the economizer outlet. It is determined by known

pressure, temperature and dryness fraction i.e. h05 = f ( P05 , T05 , χ 05 ) . The ratio

of steam condenser load to heat recovery steam generator load should range

from 45 to 50%. This range exceeds 50% if the pressure drop and the

irreversible losses were high, and the power plant operates under harsh

conditions (Molqy, 2000).

4.3.6. Hot Side Heat Recovery Steam Generator Mathematical Model

Assuming the flue gases to behave as ideal gas the heat recovery steam

generator load (kW) can be written as:

Q HRSG = m& C , total ε HRSG C P , g (T014 − T013 ) (4.68)

The total mass flow rate of cooled gas turbine m& C , total across the HRSG is

determined by Eq. (4.21). The heat recovery steam generator effectiveness

ε HRSG is taken to be 0.90. The specific heat of flue gases via the recovery C P , g

is evaluated by Eq. (4.16).

Heat recovery steam generator inlet temperature T014 is estimated as

follows: T014 = T04 − ΔT04 − 014 , where ΔT04 − 014 is the temperature drop from the

gas turbine exit to the HRSG inlet. It is taken to be about 3 °C.

Flue gases temperature (K) at point 18, 19 and 13 (see Fig. 4.5 and 4.6) can be

written respectively as:

QSH
T018 = T014 − (4.69)
ε HRSG C P , g m& C , total

QSG
T019 = T018 − (4.70)
ε HRSG C P , g m& C , total

Q ECON
T013 = T019 − (4.71)
ε HRSG C P , g m& C , total

It should be noted that in the above equations the specific heat C P , g was

calculated at the corresponding average (inlet-outlet) temperature.

Recovery exit temperature T013 is also calculated as T013 = TECON , in + PPECON

(Molqy, 1996). Here TECON , in is the economizer inlet temperature (K) (state 11)
‫أداء اﻟﻤﻨﺸﺄة اﻟﻤﺸﺘﺮآﺔ )ﻋﻨﻔﺔ ﻏﺎزﻳﺔ – ﺑﺨﺎرﻳﺔ( ﻣﻊ ﺗﺒﺮﻳﺪ اﻟﻬﻮاء اﻟﺪاﺧﻞ إﻟﻲ اﻟﻀﺎﻏﻂ‬
‫ﺑﺎﺳﺘﺨﺪام اﻟﺘﺒﺮﻳﺪ اﻻﻣﺘﺼﺎﺻﻲ اﻟﻌﺎﻣﻞ ﻋﻠﻰ ﺣﺮارة اﻟﻐﺎز اﻟﻌﺎدم‬

‫إﻋﺪاد‬
‫ﻋﻼء ﺣﺴﻴﻦ ﻋﺒﺪ اﷲ ﺟﺒﺮ‬

‫اﻟﻤﺸﺮف‬
‫أ‪ .‬اﻟﺪآﺘﻮر ﻳﻮﺳﻒ زرﻳﻘﺎت‬

‫اﻟﻤﺸﺮف اﻟﻤﺸﺎرك‬
‫أ‪ .‬اﻟﺪآﺘﻮر ﻳﻮﺳﻒ اﻟﻨﺠﺎر‬

‫ﻣﻠﺨﺺ‬

‫ﺗﻨﺎوﻟ ﺖ ه ﺬة اﻟﺪراﺳ ﺔ ﻋﻤﻠﻴﺘ ﻲ اﻟﻨﻤﺬﺟ ﺔ و اﻟﻤﺤﺎآ ﺎة ﻟﻤﻨﺸ ﺄﺗﻲ ﺗﻮﻟﻴ ﺪ اﻟﻄﺎﻗ ﺔ اﻟﻜﻬﺮﺑﺎﺋﻴ ﺔ اﻟﻤﺸ ﺘﺮآﺔ‬

‫)ﻋﻨﻔ ﺔ ﻏﺎزﻳ ﺔ – ﺑﺨﺎرﻳ ﺔ( و)ﻋﻨﻔ ﺔ ﻏﺎزﻳ ﺔ – ﻣﺴ ﺘﺮﺟﻊ ﺣ ﺮاري( ﻣ ﻊ ﺗﺒﺮﻳ ﺪ اﻟﻬ ﻮاء اﻟ ﺪاﺧﻞ اﻟ ﻰ‬

‫اﻟﻀﺎﻏﻂ‪.‬‬

‫ﻋﺪة ﻋﻤﻠﻴﺎت ﺗﻜﻨﻮﻟﻮﺟﻴﺔ ﺗ ﻢ ﺗﻄﺒﻴﻘﻬ ﺎ ﻟﺘﺒﺮﻳ ﺪ اﻟﻬ ﻮاء اﻟ ﺪاﺧﻞ اﻟ ﻰ اﻟﻀ ﺎﻏﻂ ﻓ ﻲ آﻠﺘ ﺎ اﻟﻤﻨﺸ ﺄﺗﻴﻦ‪ ,‬ﻣﺜ ﻞ‬

‫اﻟﺘﺒﺮﻳ ﺪ ﺑﺎﺳ ﺘﺨﺪام اﻷﺛ ﺮ اﻻﻣﺘﺼﺎﺻ ﻲ )أﻣﻮﻧﻴ ﺎ ﻣ ﻊ ﻣ ﺎء( و)ﺑﺮوﻣﻴ ﺪ اﻟﻠﻴﺜ ﻮم ﻣ ﻊ ﻣ ﺎء( ﺑﺎﻻﺿ ﺎﻓﺔ إﻟ ﻲ‬

‫اﻟﺘﺒﺮﻳﺪ اﻟﺒﺨﺎري وأﻳﻀﺎ ﺗﻘﻨﻴﺘﻲ اﻟﺘﺮﻃﻴﺐ ﻣﻌﺘﺪﻟﺔ وﻣﺮﺗﻔﻌﺔ اﻟﻀﻐﻂ‪ .‬ﻟﻘﺪ ﺗ ﻢ إﺳ ﺘﻨﺰاف ﺟ ﺰء ﻣ ﻦ اﻟﺒﺨ ﺎر‬

‫ﻓ ﻲ ﻣﻨﺸ ﺄة اﻟﺘﻮﻟﻴ ﺪ اﻟﻤﺸ ﺘﺮآﺔ )ﻋﻨﻔ ﺔ ﻏﺎزﻳ ﺔ – ﺑﺨﺎرﻳ ﺔ( ﻟﺘﻐﺬﻳ ﺔ ﻣﻮﻟ ﺪ ﺁﻟ ﺔ اﻟﺘﺒﺮﻳ ﺪ اﻻﻣﺘﺼﺎﺻ ﻴﺔ ﺑﻴﻨﻤ ﺎ‬

‫أﺟﺮﻳﺖ ﻋﻤﻠﻴﺔ اﻟﺘﺴﺨﻴﻦ ﻟﻠﻤﻮﻟﺪ ذاﺗ ﻪ ﺑﻮاﺳ ﻄﺔ ﺣ ﺮارة اﻟﻐ ﺎزات اﻟﻌﺎدﻣ ﺔ ﻓ ﻲ اﻟﻤﻨﺸ ﺄة اﻟﻤﺸ ﺘﺮآﺔ )ﻋﻨﻔ ﺔ‬

‫ﻏﺎزﻳﺔ – ﻣﺴﺘﺮﺟﻊ ﺣﺮاري( ﺑﺎﺳﺘﺨﺪام ﺑﺨﺎرﻣﺎء آﻮﺳﻴﻂ ﺗﺴﺨﻴﻦ‪.‬‬

‫إن اﻟﺒﺮﻧﺎﻣﺞ اﻟﻤﺤﻮﺳﺐ )ﻋﻤﻠﻴﺘﻲ اﻟﻨﻤﺬﺟﺔ و اﻟﻤﺤﺎآﺎة( اﻟﺬي ﺗﻢ إﻋﺪادﻩ ﻓ ﻲ ه ﺬة اﻟﺪراﺳ ﺔ ﻗ ﺪ أﺧﻀ ﻊ‬

‫ﻟﻌﺪة إﺧﺘﺒﺎرات وذﻟﻚ ﻟﻤﻌﺮﻓﺔ ﻣﺪى دﻗﺔ هﺬا اﻟﻨﻈﺎم اﻟﻤﻄﻮر‪.‬ﻟﻘﺪ ﺗ ﻢ اﻟﺘﺄآﻴ ﺪ ﻋﻠ ﻰ ذﻟ ﻚ ﻣ ﻦ ﺧ ﻼل إدﺧ ﺎل‬
‫ﻋ ﺪة ﺑﻴﺎﻧ ﺎت ﺗﻘﻨﻴ ﺔ ﻣﺘﻌﻠﻘ ﺔ ﺑﺎﻟﻌﻨﻔ ﺎت اﻟﻐﺎزﻳ ﺔ اﻟﺼ ﺎدرة ﻋ ﻦ أﺷﻬﺮاﻟﺸ ﺮآﺎت اﻟﺼ ﺎﻧﻌﺔ وﻣﻨﺸ ﺂت اﻟﺘﻮﻟﻴ ﺪ‬

‫اﻟﻤﺸ ﺘﺮآﺔ )ﻋﻨﻔ ﺔ ﻏﺎزﻳ ﺔ – ﺑﺨﺎرﻳ ﺔ( إﻟ ﻰ ه ﺬا اﻟﺒﺮﻧ ﺎﻣﺞ وآﺎﻧ ﺖ ﻧﺴ ﺒﺔ اﻟﺨﻄ ﺄ ﻗﺮﻳﺒ ﺔ إﻟ ﻰ اﻟﺼ ﻔﺮ‪ .‬إن‬

‫اﻟﺒﻴﺎﻧﺎت اﻟﻤﺪﺧﻠﺔ ﻓ ﻲ ﺣﺎﻟ ﺔ اﻟﺘﻮﻟﻴ ﺪ اﻟﻤﺸ ﺘﺮك )ﻋﻨﻔ ﺔ ﻏﺎزﻳ ﺔ – ﺑﺨﺎرﻳ ﺔ( ﻗ ﺪ أﺧ ﺬت ﻣ ﻦ ﻣﺤﻄ ﺔ ﺗﻘ ﻊ ﻓ ﻲ‬

‫وﻻﻳﺔ ﻣﻮﻧﺘﺮي واﻷﺧﺮى ﻓﻲ ﻣﺎﺳﺎﺷﻮﺳﺘﺲ‪.‬‬

‫ﺗﻤ ﺖ أﻳﻀ ﺎ دراﺳ ﺔ اﻟﻨﺎﺣﻴ ﺔ اﻹﻗﺘﺼ ﺎدﻳﺔ ﻣ ﻦ ﺗﻜ ﺎﻟﻴﻒ و ﻓﺘ ﺮة إﺳ ﺘﺮﺟﺎع ﻟ ﺮأس اﻟﻤ ﺎل اﻟﻤﻮﻇ ﻒ‬

‫ﺑﺎﻹﺿ ﺎﻓﺔ إﻟ ﻰ اﻟﻘﻴﻤ ﺔ اﻟﺤﺎﻟﻴ ﺔ اﻟﺼ ﺎﻓﻴﺔ ﻟﻌ ﺪد ﻣﺨﺘﻠ ﻒ ﻣ ﻦ ﻣﻌ ﺎﻣﻼت اﻟﺘﺤﻮﻳ ﻞ اﻟﻨﻈ ﺎﻣﻲ ﻟﻠﻤﺼ ﺎرﻳﻒ‬

‫اﻟﻤﺨﺘﻠﻔﺔ زﻣﻨﻴﺎ‪ .‬ﻣﻌﺪل اﻟﺤﺴﻮﻣﺎت و اﻟﺘﻀﺨﻢ اﻟﻤﺎﻟﻲ ﺗﻢ أﺧﺬﻩ ﻣﺴﺎوﻳﺎ ‪ 0.08‬و‪ 0.03‬ﻋﻠﻰ اﻟﺘﻮاﻟﻰ‪.‬‬

‫إن اﻟﻨﺘ ﺎﺋﺞ اﻟﻨﻬﺎﺋﻴ ﺔ أﺷ ﺎرت إﻟ ﻰ أن ﻣ ﺮدود اﻟﻌﻨﻔ ﺔ اﻟﻐﺎزﻳ ﺔ اﻟﻌﺎﻣﻠ ﺔ ﺑﺈﺳ ﺘﻄﺎﻋﺔ ‪ 203 MW‬وﻓﻘ ﺎ‬

‫ﻟﻠﺸﺮوط اﻟﻨﻈﺎﻣﻴﺔ ﻳﺴﺎوى ﺗﻘﺮﻳﺒ ﺎ ‪ 50%‬وذﻟ ﻚ ﺑﺈﺳ ﺘﺨﺪام ﺗﻘﻨﻴﺘ ﻲ اﻟﺘﺒﺮﻳ ﺪ ﺑﺎﻟﺘﺮﻃﻴ ﺐ ﻣﻌﺘﺪﻟ ﺔ أوﻣﺮﺗﻔﻌ ﺔ‬

‫اﻟﻀﻐﻂ ﻋﻨﺪ ﻣﺪﺧﻞ اﻟﻀﺎﻏﻂ‪ .‬إن ﻣﺮدود ﻣﻨﺸﺄة اﻟﺘﻮﻟﻴﺪ اﻟﻤﺸﺘﺮآﺔ )ﻋﻨﻔ ﺔ ﻏﺎزﻳ ﺔ – ﻣﺴ ﺘﺮﺟﻊ ﺣ ﺮاري(‬

‫و )ﻋﻨﻔﺔ ﻏﺎزﻳﺔ – ﺑﺨﺎرﻳﺔ( آﺎن ﻣﺴﺎوﻳﺎ ﺗﻘﺮﻳﺒﺎ إﻟ ﻰ ‪ 56%‬و ‪ 60%‬ﻋﻠ ﻰ اﻟﺘ ﻮاﻟﻰ وذﻟ ﻚ ﺑﺈﺳ ﺘﺨﺪام ﺁﻟ ﺔ‬

‫اﻟﺘﺒﺮﻳﺪ اﻻﻣﺘﺼﺎﺻﻴﺔ أو اﻟﺒﺨﺎرﻳﺔ‪.‬‬