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An applicable method for gas turbine efficiency improvement. Case study: Montazar Ghaem power plant, Iran
Journal of Natural Gas Science and Engineering 28 (2016) 95e105
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Journal of Natural Gas Science and Engineering journal homepage: www.elsevier.com/locate/jngse
An applicable method for gas turbine efficiency improvement. Case study: Montazar Ghaem power plant, Iran Afsaneh Noroozian, Mokhtar Bidi* Faculty of Mechanical and Energy Engineering, Shahid Beheshti University, A.C., Tehran, Iran
a r t i c l e i n f o
a b s t r a c t
Article history: Received 5 September 2015 Received in revised form 19 October 2015 Accepted 16 November 2015 Available online 1 December 2015
Compressor inlet air cooling is one of the most well-known methods to improve gas turbine power plant performance. In this study, pressure reduction valve in natural gas pressure reduction station is replaced by a turbo-expander to use the potential energy of the compressed gas. The turbo-expander shaft is connected to a mechanical chiller to produce refrigeration; the produced refrigeration is used to decrease the compressor inlet air temperature. Moreover, thermodynamic and exergetic analyses are carried out and the effect of compressor air cooling on the performance of the plant is studied. To do so, Montazer Ghaem power plant is considered as the case study. Results showed that using cooling system causes 3.2% temperature drop which leads to 1.138% increment in both thermal efficiency and net output power in the warmest month. Exergetic analysis reveals that using the cooling system leads to a higher exergy efficiency and hence lower exergy destruction. Also combustion chamber with 81.26 MW has the highest amount of exergy destruction which decreases to 77.36 MW after implementation of the cooling system. In January, exergetic efficiency of total plant has 1.86% enhancement and exergy destruction reduces about 3.8 MW. © 2015 Elsevier B.V. All rights reserved.
Keywords: Gas turbine power plant Turbo-expander Compressor inlet air cooling system Efficiency improvement Exergetic analysis
1. Introduction The optimization and reduction of energy consumption has an important role in the contemporary world. Increasing the efficiency of various devices and decreasing their consumed energy is preferred. Gas turbine is a device which has various applications such as producing electricity in power plants, especially those which require less commissioning time. The efficiency and output power of gas turbines are related to many different parameters. The aim of this study is to investigate the influence of ambient temperature on gas turbine plants output power. One Celsius degree temperature increment leads to approximately one percent reduction of the gas turbine rated capacity (Mohanty and Paloso, 1995). There are as many methods to reduce compressor inlet air temperature so as to improve the gas turbine performance. The most important ones are as follows (Kakaras et al., 2006; Hosseini et al., 2007; Chacartegui et al., 2008; Ehyaei et al., 2011):
* Corresponding author. E-mail address: [email protected] (M. Bidi). http://dx.doi.org/10.1016/j.jngse.2015.11.032 1875-5100/© 2015 Elsevier B.V. All rights reserved.
a) Mechanical chiller or absorption chiller: mechanical chiller cools the inlet air so its temperature drops to below the wet bulb temperature. Mechanical chillers are driven by electric motors, steam turbines or engines. These type of chillers use a classical mechanical compression cycle. The inlet air is passed across the cooling coils, through which intermediate refrigerant fluid is circulated. Absorption chillers require steam or hot water as the basic source of energy and they use absorption refrigeration cycle with the working fluids of water/ammonia or lithium bromide/water. Absorption chillers also require less electric energy than mechanical chillers. The main difference of these two types of chillers is in the way that the refrigerant is changed into liquid. This is the reason why absorption chillers do not have any moving parts. b) Chilled water storage or ice harvesting: chilled water storage and ice storage, store sensible and latent heat energy respectively. Ice storage is mostly profitable for small storage capacity. c) Media evaporative cooling and Fogging: the evaporative cooler is a wetted rigid media. As the water distributes from the top, hot air passes through the media and cause water evaporation. So, the air is cooled during the process and then enters the compressor. Fogging is another evaporative cooling technology in which the fine droplets of water are added to the inlet air by
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means of high pressure nozzles. Water consumption in fogging system is more than that of evaporative cooling, but fogging system has higher effectiveness. In a method suggested by Mohanty and Paloso (1995), the heat of flue gas was directly used by an absorption chiller to provide cooling system. Therefore the gas turbine output power increases, the waste heat is recovered and the environmental impacts are decreased. They also concluded that 11% increase in plant's energy output could be achieved. M. De Lucia et al. (1995) discussed differences between absorption and evaporative cooling systems. They concluded that, although evaporative cooling systems have a lower cost, but absorption ones are much more effective especially for base load production plants. Combination of these two cooling systems which has the highest benefit was likewise. Badran (1999) worked on the influence of parameters which could affect the gas turbine performance, such as compressor inlet temperature and turbine inlet temperature. He emphasized that as the compressor inlet air temperature increases, thermal efficiency decreases and also ambient condition would have less effect on efficiency by choosing the best location for power stations. M. Chaker et al. (2003) have studied available direct evaporative cooling hours of 122 locations in the U.S.A with inlet fogging system for gas turbines. Their focus was mostly on climate conditions. Ameri and Hejazi (2004) suggested an air cooling system that uses a steam absorption chiller in Chabahar (Iran) power plant. They concluded that this system with payback period of 4.2 years, would have led to 11.3% increase in average output power. Kakaras et al. (2006) have compared different air cooling methods and their integration. They mentioned that investment costs of evaporative cooling technology are lower than other methods, while absorption chiller system has much higher capacity for increasing electricity generation. Hosseini et al. (2007) evaluated the influence of media evaporative cooling system on gas turbines performance in a combined cycle power plant. Their results showed that this cooling method causes 19 C reduction in inlet air temperature which leads to 11 MW growth in gas turbine output power with a payback period of about 4 years. Shi et al. (2010) performed a research aimed at improvement of combined cycle power plant. They demonstrated that with integration of inlet air cooling, inter cooling and LNG gasification within a combined cycle power station, there would be an increment in net electrical efficiency and output power. De Sa and Al Zubaidy (2011) focused on the relation between ambient temperature and gas turbine efficiency and output power. They mentioned that with increase of temperature, efficiency and output power are reduced. They also established an empirical relationship between ambient humidity and system performance. In addition to conventional cooling methods, there are modern cooling technologies such as ceramic tubes membrane (Zeitoun et al., 2014). Zeitoum et al. worked on an experimental sample and demonstrated that by having both latent and sensible heat transfer, in this model the air passes over the ceramic tube matrix and the water runs through the ceramic tubes. So the ambient temperature decreases and the relative humidity increases with no erosion in compressor blades. There are other methods for gas turbine performance enhancement. Using liquid nitrogen spray is one way to reduce turbine inlet air temperature in Integrated Gasification Combined Cycles (Morini et al., 2015). Some advantages of this method are that it does not need a large amount of water despite common thermal storage systems and there won't be pressure losses at the compressor inlet. Oyedepo and Kilanko (2012) used evaporative cooler for performance enhancement of a gas turbine power plant. They mentioned that 5 C drop in compressor inlet air temperature, causes 5e10% increment in net output power and 2e5% thermal efficiency enhancement. They also showed the positive effect of
pressure ratio on their results. Ibrahim et al. (2011) showed that if compressor inlet air temperature increase 1 C, the gas turbine power output drop about 1%. They also showed that by increasing ambient air temperature, air mass flow rate reduces and finally cycle efficiency decreases. Gas turbine power plants performance in warm and relatively dry climate is very important, so cooling compressor inlet air in these conditions is the topic of many researches (Tobi, 2009; Abam et al., 2012; Jarze˛ bowski et al., 2012). Zaki et al. (2011) demonstrated that by the use of evaporative cooler in hot humid climate, there would be limitation in gas turbine capacity improvement (not more than 5%e7%). Because compressor inlet air temperature should not be less than wet bulb temperature. Farzaneh-Gord and Deymi-Dashtebayaz (2009, 2011) used a similar method for inlet air cooling system to enhance the output power, but they directly used expansion turbine outlet flow for cooling the compressor inlet air by the use of a heat exchanger. Implementation of their model is not possible for a constructed power plant, because their method causes fundamental changes in fuel supply system. As a result, installing of this model needs a long overhaul of power plant and this is against the power plant owners' expectations. On the other hand, because of long distance between pressure reduction station and gas turbine power plant, there would be a possibility to phase change, so this project is very hard to execute. Moreover there were two fundamental shortcomings in their thermodynamic modeling which leads to less accurate results. The first one is that they did not correctly relate the inlet air cooling to output power, another problem is that they used ideal Brayton assumption in their calculations, therefore there will be miscalculations and it could not be used in real industrial decisions. For example Fig. 1 shows differences between the results of simulation procedure with ideal Brayton cycle in comparison with real data for a 134.140 MW gas turbine model (GE, model MS9001E with 34.6% efficiency at shaft output). There is a 7.92% discrepancy between efficiency of ideal Brayton cycle with the catalogue data, while there is no considerable error in the proposed method. In the present study an applicable method for an installed power plant is provided. In the proposed model, a mechanical chiller is used for reducing compressor inlet air temperature which its required power is obtained from a turbo-expander. The turboexpander is replaced in pressure reduction station with a pressure reduction valve. Using a turbo-expander is an appropriate way to recover pressure energy of natural gas in pressure reduction stations (Farzaneh-Gord et al., 2015). Fuel pressure must be reduced from 5e7 MPa to 1.5e4.0 MPa, in the natural gas pressure reduction station (Po zivil, 2004). Andrei et al. (2014) demonstrated that by the use of turbo-expanders instead of throttle valves or pressure reducing valves in pressure reducing stations of the whole country, the sum of power generation could be significant. In this study a preheater is used before the expansion turbine to prevent from production of liquid or solid phase at the discharging of the station. This article compares basic model with proposed model, from the viewpoints of thermodynamics and exergy. As discussed earlier, in previous studies ideal Brayton cycle was used for simulation which caused miscalculations as mentioned in Fig. 1. So in this paper a real cycle is simulated for more accuracy. The proposed model is an applicable method to use potential energy of natural gas unlike previous studies which were impracticable. It is for the first time that exergetic analysis of each component and whole cycle is done for a real gas cycle with cooling system. Exergy destruction rate and exergy efficiency of different points in the cycle are also calculated. Other power plants can use the methodology of the proposed model to enhance the plant power and efficiency, because this model is applicable and does not involve executive problems unlike previous models. Effect of different amounts of COP on net output
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Fig. 1. Comparison between results of ideal Brayton cycle and the real gas cycle for GE MS9001E gas turbine.
power is also represented in the results section, so one can see the effects of different condition of cooling system on the output of gas turbine power plants. An upward trend in thermal efficiency, net output power, exergetic efficiency and exergy destruction by the use of proposed model is also predicted for other power plants. However to obtain accurate and applicable results for other power plants, the similar analyses is required using the design data of the power plants and its environmental conditions.
3. Thermal modeling 3.1. Basic model The compressor inlet temperature is equal to ambient temperature in the basic model which neglects the cooling effect and simulates the cycle under ambient conditions.
T1;air;b ¼ Tamb + h1;b ; s1;b P1;air;b ¼ Pamb
2. System description Two different models including basic model and proposed model are investigated in this paper. The basic model consists of a preheater, pressure reduction valve, combustion chamber, compressor and gas turbine as shown in Fig. 2, While Fig. 3 shows the proposed model which has an expansion turbine instead of a pressure reduction valve and it also consists of some auxiliary equipments like heat exchanger and mechanical chiller. Calculation procedures of these two cases are as follows. The flow chart of the basic and proposed model analyses is shown in Fig. 4.
(1)
In this model, inlet air is considered to be an ideal gas and its density is calculated according to the following equation:
r1;air;b ¼
MP1;air;b Ru T1;air;b
(2)
Where M is the molar mass and Ru is the universal gas constant. Volume flow rate of the compressor can be calculated as follow:
Fig. 2. A schematic diagram of the basic model.
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Fig. 3. A schematic diagram of the proposed model.
m_ air;b V_ c;b ¼ r1;air;b
(3)
Wherem_ air;b , is the air mass flow rate of the basic system. The pressure of the air which leaves the compressor (P2,air,b) is specified by the following equation and the temperature of the outlet air (T2,air,b) is obtained from ideal gas equation, considering the value of 0.85 for the isentropic efficiency of the compressor (hc):
P2;air;b ¼ P1;air;b r + 2 3 g1 ! h2;b ; s2;b g T P 5 þ T1;air;b T2;air;b ¼ 1;air;b 4 2;air;b 1 hc P1;air;b
(4)
r is the compression ratio and g is the ratio of specific heat which is calculated using the following formula:
g ¼ Cp=Cv
(5)
Where Cp and Cv are specific heat at constant pressure and volume, respectively. Both of them are calculated at the temperature of T1,air,b. The compressor inlet work is also estimated from:
_ _ air;b h2;b h1;b W c;b ¼ m
(6)
The next step is to calculate the amount of additional heat to preheat natural gas (process5,NG,b_6,NG,b) and also the temperature of natural gas entering the combustion chamber. By applying the first law of thermodynamics:
P5;NG;b ¼ + T5;NG;b ¼ h5;b Q_ in;NG;b ¼ m_ NG;b h6;b h5;b P6;NG;b ¼ P5;NG;b h6;b T6;NG;b Fig. 4. Flow chart of calculation procedure.
(7)
Where m_ NG;b is the mass flow rate of natural gas which passes through the pressure drop station.
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P7;NG;b ¼ P6;NG;b rstation + T7;NG;b h7;b ¼ h6;b
(8)
rstation is the compression ratio of pressure drop station. Using the analysis of the combustion process between the fuel flow and air flow (7,b and 2,b), characteristics of flue gas are defined in point 3. Enthalpy of flue gas in point 3 is calculated using the analysis of the chemical reaction equation for the combustion:
h3;b ¼
X
xi hi
(9)
Analogous to the mass fraction, the mole fraction of component i (xi), is defined as the ratio of the number of moles of i, (ni), to the total number of moles in the mixture, (n):
n xi ¼ i n
(10)
Where n is the stoichiometric coefficient of the chemical equation for the combustion reaction. The heat delivered by the combustion chamber is determined from using the lower heating value of the natural gas (LHV (kj/kg)).
Q_ in;air;b ¼ m_ NG;b LHV
(11)
By assuming the use of a specific turbine, pressure and temperature of point 4 is obtained.
P4;b ¼ T4;b ¼ h4;b ; s4;b
(12)
(13)
Where m_ g;b is the mass flow rate of flue gas and it is given by:
m_ g;b ¼ m_ air;b þ m_ NG;b
(14)
Finally the thermal efficiency of the gas turbine is calculated as follows:
hth;b ¼
_ _ W T;b W c;b _ Q
(16)
So the expansion turbine power is achieved using the following relation:
_ _ NG;p h6;p h7;p W EXT;p ¼ m
(17)
The turbine shaft is connected to a mechanical chiller. Chiller removes the heat from liquid, this liquid can then be circulated through a heat exchanger to cool the air which enters the compressor. The cooling power produced by chiller is equal to:
_ EXT; p COP Ql;chiller ¼ W
(18)
Where COP is the coefficient of performance and in this project, it is considered to be 2.5. This simplification is because of prevention from modeling a refrigeration cycle which causes complicacy of calculations, so a fixed amount is assumed for COP which is suitable for a refrigeration cycle. The next step is to compute the compressor inlet air temperature. As it was mentioned earlier, temperature and pressure in point 1 is equal to ambient conditions. By applying equations related to the heat exchanger, compressor inlet temperature and pressure can be obtained. So the density of air flow is achieved.
Q_ l;chiller hH;E ¼ m_ air;p Cpair T1;air;p T1;c;p
(19)
Where hH,E is the isentropic efficiency of heat exchanger and is assumed to be 0.9. Cpairis the specific heat capacity of air.
The turbine output power is equal to:
_ _ g;b h3;b h4;b W T;b ¼ m
T7;p ¼ P7;p ¼ h7;p ; s7;p
99
(15)
in;air;b
_ _ In the above equation, the term (W T;b W c;b ) shows the output _ power of the whole cycle (W net ).
3.2. Proposed model In the proposed model, some new equipments were added to the cycle, so previous relations should be replaced by new ones. Note that Cp and Cv in the proposed model are calculated at the temperature of T1,c,p. The effect of relative humidity is neglected, because of its yearly average amount which is less than 40% at Montazar Ghaem power plant location and hence it will not cause in much miscalculations. The cooling system is simplified using a constant value for the COP. Within the heat exchanger, the specific heat of air is supposed to be constant. Other model assumptions are listed in Table 2. These quantities are based on Montazar Ghaem's data. By assuming the use of a specific expansion turbine instead of a pressure reduction valve, pressure and temperature of point 7 is obtained.
T1;c;p ¼ P1;c;p M P1;c;p ¼ r1;c;p ¼ R T u 1;c;p
(20)
In the case of installing inlet air cooling system, the inlet compressor pressure is assumed to be unchanged (same as ambient pressure). According to Eq. (20), with decreasing the compressor inlet temperature, the density of air flow similarly, will reduce. Because the volume flow rate of the compressor in both models is constant, so air mass flow rate of proposed model (m_ air;p ) will increase. Here, flow characteristics in different parts of the proposed system are calculated similar to the basic model. By changing flow conditions in compressor and combustion chamber, Eqs. (6), (11), (13)e(15), are rewritten as follows:
_ c;p ¼ m_ W air;p h2;p h1;p
(21)
Q_ in;air;p ¼ m_ NG;p LHV
(22)
_ _ g;p h3;p h4;p W T;p ¼ m
(23)
Where m_ g;p is the mass flow rate of flue gas and it is given by:
m_ g;p ¼ m_ air;p þ m_ NG;p
hth;p ¼
_ _ W T;p W c;p _ Q in;air;p
(24)
(25)
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3.3. Exergy balance equation
Exergy is the maximum amount of useful work which can be obtained from a machine. Exergy won't be stored during a process, but it destroys because of irreversibility. Each single component of a cycle can be analyzed individually in this method and each component's contribution to exergy's waste in the whole system could be obtained. An exergetic analysis was derived for each component of GE-MS9001 gas turbine power plant. Exergetic efficiency and exergy destruction of cycle with compressor inlet air cooling system and without it were compared. In this part, equations are written for an ideal gas having a constant specific heat. The thermo-mechanical exergy flow can be divided into two branches, thermal and mechanical exergy flows (Al-Doori, 2012), which are related to the temperature and pressure of the stream. Chemical exergy is related to the composition of the fuel stream. The general exergy balance equation is as follows (Ebadi and GorjiBandpy, 2005): ch E_ þ
X
T E_ in
out w
! T E_ out
out
in
X
X
. S_out þ Q_ cv T ref
þ
!
X
M E_ in
X
! M E_ out
out
in
þ Tref
X
S_in
in
(26) In this equation, subscripts in and out respectively show the inlet and outlet of the exergy flow through each component. Subscript ref indicates reference environmental conditions which are:Pref ¼ 100 (kp), Tref ¼ 25 ( C). Ru is universal gas constant and equals to 0.287 (kj/kg.k).Q_ cv shows the heat transfer between component and environment. P P _ The term Tref ð S_in Sout Þ shows exergy destruction rate out in (E_ D ). The thermal and mechanical exergy flow rates can be written as follows (Ebadi and Gorji-Bandpy, 2005):
"
T Tref
M E_ ¼ m_ R Tref ln
P Pref
Tref
T ln Tref
!# (27)
! (28)
For chemical exergy, following approximation is used (Ahmadi et al., 2011): ch E_ ¼ x LHV m_ NG 1000
(29)
=
Where x is chemical exergy/energy ratio and in this study, its value for CH4 is 1.06.
3.3.1. Exergy balance and exergy efficiency equations for components Exergy balance and exergy efficiency equations for each component are as follows: (Al-Doori, 2012): Compressor:
M T T M _c E_ 1 E_ 2 þ E_ 1 E_ 2 þ Tref S_1 S_2 ¼ W
(30)
_c W
(31)
Combustion chamber:
M T ch T T M M E_ þ E_ 2 þ E_ 7 E_ 3 þ E_ 2 þ E_ 7 E_ 3 þ Tref S_2 þ S_7 S_3 . þ Q_ cv T ref
!
¼0 (32)
hex;cc
T M E_ 3 þ E_ 3 ¼ T T M M ch E_ 2 þ E_ 2 þ E_ 7 þ E_ 7 þ E_
(33)
Gas turbine:
T M T M _T E_ 3 E_ 4 þ E_ 3 E_ 4 þ Tref S_3 S_4 ¼ W
hex;c ¼
¼ E_
T E_ ¼ m_ Cp
hex;c ¼
T T M M E_ 2 þ E_ 2 E_ 1 þ E_ 2
T E_ 3
þ
M E_ 3
_ W T T M E_ 4 þ E_ 4
(34)
(35)
The same equations could be used to calculate exergy performance of the system with cooling system. 4. Case study Montazar Ghaem power plant which is located at 7 km Mallard road in Alborz province, consists of different power generator plants as follows: Four steam power generators, six gas turbine generators and three combined cycle generators. This plant has a total capacity of 625.2 MW. Each combined cycle power unit includes two GEMs9001E gas turbines, with nominal capacity of 110 MW. The technical data of these turbines are shown in Table 1. In this study, the validity of proposed model will be checked by Montazar Ghaem's data. Table 2 shows the recorded values in Montazar Ghaem plant. Table 3 shows average monthly temperatures of Alborz province where the power plant is located. 5. Results and discussions For verification of the developed code, the modeling results were compared with the obtained results of Ameri and Hejazi (2004) and a good agreement between these two studies were found. One of Montazar Ghaem's gas units was selected as case study for thermal and exergetic analysis. In this section, the effects of inlet air cooling on gas turbine performance are investigated. As discussed earlier, the proposed model is implemented on the GEMS9001E gas turbine which is installed within an existing power plant in Iran. Figs. 5e10 show the effects of temperature drop on gas cycle characteristic parameters. By the use of a mechanical chiller coupled with a heat exchanger, the temperature of compressor inlet air can be reduced as shown in Fig. 5. According to Eq. (17) and Eq. (18), cooling power depends on the coefficient of performance (COP) and reduction station inlet and outlet conditions. Therefore by assuming a
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Table 1 Feature of gas turbine engine used for simulation (ISO conditions, natural gas) (GE Gas turbines, 2006). Ms9001E
ISO rated power (kW)
Heat rate (kJ/kWh)
Efficiency (%)
Pressure ratio
Exhaust temperature ( C)
Exhaust flow (kg/s)
134,140
10,397
34.6
12.7
540
410.0
Table 2 Montazar Ghaem's data which is taken from its official Website. Description
Recorded value
Unit
Mass flow rate of air, entering the compressor (m_ air ) Mass flow rate of natural gas through of the pressure drop station (m_ NG ) compression pressure ratio (r) Ambient temperature (Tamb) Ambient (compressor inlet) pressure (P1,air) Inlet pressure of the pressure reduction station (P5,NG) Outlet pressure of the pressure reduction station (P7,NG) Inlet temperature of the pressure reduction station (T5,NG) Outlet temperature of the pressure reduction station (T7,NG) Outlet temperature of the gas turbine (T4) Specific heat at constant pressure (Cpair)
350 4.16 10:1 25 0.9 51.7 17.2 60 11 525 1.004
(kg/s) (kg/s) e ( C) (bar) (bar) (bar) ( C) ( C) ( C) (kj/kg K)
The analysis carried out by taking the weather data of Iran Meteorological Organization for Alborz province weather, into account.
Table 3 Average monthly temperature of Alborz (Iran) province ( C) which is taken from I.R. of Iran Meteorological Organization Website. Year
Jan
Feb
Mar
Apr
May
Jun
Jul
Aug
Sep
Oct
Nov
Dec
2014
3.6
3.5
10.4
16
21.8
26.3
28.9
28.7
24.8
15.1
7.7
5.9
Fig. 7. Compressor required power.
Fig. 5. Compressor inlet temperature.
Fig. 8. Gas turbine output power.
Fig. 6. Turbine air mass flow rate.
constant value for the coefficient of performance and reduction station conditions, a constant value of cooling power and a constant temperature drop during a year are obtained in the proposed model. Temperature drop is about 3.2% during a year.
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Fig. 9. Net output power of cycle.
Fig. 10. Gas turbine thermal efficiency.
Fig. 6 shows that inlet air cooling will make the air mass flow rate increase. Supposing that the air pressure is constant in Eq. (20), the air density increases by reduction of air temperature. Therefore, with this supposition that the air volume flow rate (V_ c ) is constant, the air mass flow rate (m_ air ) in the proposed model will increase. The air mass flow rate in warm seasons is less than cold seasons because air density decreases in warm seasons. The percent of increase in the air mass flow rate due to the use of proposed model is about 3.39 in July and 3.52 in February. As it is shown in Figs. 7e8 it's clear that the influence of temperature drop on gas turbine power is more than its influence on compressor required power, but for both of them, changes are not tangible. Change in compressor inlet air temperature, causes alteration in specific enthalpy. As is construed from Eq. (4), any changes in the specific heat ratios and compressor inlet temperature leads to variation in temperature and specific enthalpy of point 2 rather than its value in basic model. Variation of specific heat ratios (g) is due to different quantities of Cp and Cv in the proposed model and the variation of Cp and Cv is related to compressor inlet temperature drop. Finally, according to Eq. (21), in the model with cooling system, enthalpy difference decreases more than the air mass flow rate increment. As a result the required power of compressor decreases a bit. Reduction of compressor required power in warmer months like July (0.18%) is a bit less than cold months (0.26% in February). Since the mass flow rate of flue gas entering the gas turbine is the sum of air mass flow rate and natural gas mass flow rate, with the increase of air mass flow rate and constant amount of natural gas mass flow rate, flue gas mass flow rate will increase. By the use of a special gas turbine, thermodynamic parameters of point 4 remain constant in both models. Specific enthalpy of point 3 is obtained from combustion analysis and is fixed in both models.
According to Eq. (13), the amount of gas turbine output power is multiplication of enthalpy difference of points 3 to 4 and flue gas mass flow rate. In the proposed model, the enthalpy difference is unchanged and as the flue gas mass flow rate increases, the gas turbine output power also increases. Rate of changes in July and February are 0.45% and 0.33% respectively. The net output power and thermal efficiency of gas turbine cycle increase with the decrease of compressor inlet temperature, as depicted in Figs. 9e10 respectively. The net output power is the difference between gas turbine output power and compressor required power. So as discussed earlier, in the proposed model, gas turbine output power will increase till the compressor required power decreases and finally net output power increases subsequently. As shown in Fig. 9. Net output power is increased by 1.138% and 0.98% for July and February respectively. Since there are 6 gas units in this power plant, therefore a total increase in power will be 6484.23 kW in July and 5975.88 kW in February. Significant increment in thermal efficiency of proposed model can be explained by Eq. (25). Thermal efficiency is related to the net output power and the heat delivered by the combustion chamber. The heat delivered by the combustion chamber (Q_ in;air ) is determined from multiplication of the lower heating value of the natural gas (LHV) and its mass flow rate (m_ NG ), so with unchanged amount of these parameters in both models, the amount of delivered heat remains constant. With the increase of net output power in the model with cooling system, thermal efficiency will improve. The amounts of efficiency alterations are as same as net output power alterations, so the percentages of changes are 1.138% and 0.98% in July and February respectively. Fig. 11 shows the net output power quantities in basic and proposed model compare with real measured data of the power plant. The differences between the real data and results of the basic model show the simulation errors. It is clear that the cycle with cooling system proposed more net output power than the others. Deviations from real quantities in warm seasons are more than cold seasons especially in July and August, as depicted in Fig. 11. Effect of different amounts of COP on the net output power is shown in Fig. 12. It is clear that with decrease of COP in proposed model, net output power will be decreased. According to Eq. (18), the cooling power produced by chiller (Q_ l;chiller ) is proportional to the COP, so with lowering the COP, Q_ l;chiller will also decrease and its positive effects on the gas cycle parameters will reduce. Optimizing the cooling system parameters like COP, can be done using optimization methods such as genetic or PSO algorithms with selection of an objective function such as maximization of exergetic efficiency or net output power. To do this an accurate modeling of the cooling system is required and is not the main objective of the present study.
Fig. 11. Net output power of basic and proposed models in comparison with real quantities.
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Table 5 Net exergy flow rates and exergy destruction in the gas turbine power plant at rated conditions (Without cooling system). Component
_ ðMWÞ W
ch E_ ðMWÞ
T E_ ðMWÞ
M E_ ðMWÞ
E_ D ðMWÞ
Compressor Combustion chamber Gas turbine Total cycle
61.604 0 117.940 56.336
0 220.5 0 220.5
28.53 139.3 119.9 47.93
51.35 0.063 52.30 1.013
18.276 80.763 54.26 153.3
Table 6 Property values and thermal, mechanical and chemical exergy flow rates at various state points in the gas turbine cycle at rate conditions (With cooling system).
Fig. 12. Effect of COP on net output power of proposed model.
5.1. Exergetic analysis results Tables 4 and 6, show property values at various points in the cycle, with consideration of cooling system and without it. These parameters were measured in previous section. Tables 5 and 7, show net exergy flow rates and exergy destruction of control volumes that have been assumed around each component, with consideration of cooling system and without it. For exergetic analysis, data of one month, January, is used. The values which are shown in Tables 4 and 6 are obtained using the thermodynamic properties of each point such as pressure, temperature and mass flow rate. The analysis of different exergy streams in a system can be done using the thermophysical figures of JANAF Tables (Ebadi and Gorji-Bandpy, 2005). There's important information in Tables 5 and 7 for the comparison of the two models. The resources which generate the exergy of the cycle are shown using negative values. Fuel also has negative exergy value. Exergy flow rate of products are shown positive (Ebadi and Gorji-Bandpy, 2005). For satisfying the exergy balance of each component and also the whole cycle, the sum of exergy flow rates of recourses, products and destruction should be zero. Fig. 13 shows exergetic efficiency enhancement by the use of cooling system. In January, temperature decreases from 276.75 to 267.4 K and this reduction causes exergy efficiency of total plant increases from 57.37% to 58.44%. It is clear that efficiency increment in combustion chamber is the most, compared with compressor and gas turbine. The exergy efficiency of combustion chamber experienced an increase from 73.07% to 74.33%. Decrease of exergy destruction in total plant by the use of cooling system is shown in Fig. 14. Exergy destruction in combustion chamber has the highest amount in comparison with compressor and gas turbine. This is because combustion chamber has greater irreversibility than other components. According to temperature reduction, exergy destruction of total plant reduced from 153.3 to 149.5 MW. Also, as observed in Fig. 14, the decline in the exergy destruction of the combustion chamber is remarkable.
State point
m_ ðkg=sÞ
T (k)
P (bar)
ch E_ ðMWÞ
T E_ ðMWÞ
M E_ ðMWÞ
1c 2 3 4 7
215.290 215.290 219.450 219.450 4.160
267.4 645.6 1405.6 798.15 284.15
1 17.9 17.9 1 17.2
0 0 0 0 220.5
0.369 26.77 169.9 49.75 0.00142
0 53.15 54.17 0 1.013
Table 7 Net exergy flow rates and exergy destruction in the gas turbine power plant at rated conditions (With cooling system). Component
_ ðMWÞ W
ch E_ ðMWÞ
T E_ ðMWÞ
M E_ ðMWÞ
E_ D ðMWÞ
Compressor Combustion chamber Gas turbine Total cycle
61.444 0 118.336 56.892
0 220.5 0 220.5
26.4 143.1 120.1 49.38
53.15 0.007 54.17 1.013
18.106 77.4 54 149.506
Fig. 13. Exergetic efficiency of components and total plant.
Table 4 Property values and thermal, mechanical and chemical exergy flow rates at various state points in the gas turbine cycle at rate conditions (Without cooling system). State point
m_ ðkg=sÞ
T (k)
P (bar)
ch E_ ðMWÞ
T E_ ðMWÞ
M E_ ðMWÞ
1 2 3 4 7
208 208 212.160 212.160 4.16
276.75 667.9 1422.869 798.15 284.15
1 17.9 17.9 1 17.2
0 0 0 0 220.5
0.1689 28.7 168 48.1 0.00142
0 51.35 52.30 0 1.013
Fig. 14. Exergy destruction in components and total system.
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6. Environmental impacts Nowadays, environmental impacts are matter of concern to industry owners. Many studies have been done about ways to reduce industrial pollutions using classical methods (Valipour et al., 2013a). Valipour et al. (2012) designed an environmental flow diagram (EFD) for determining environmental impacts of industries. They studied on air, water and soil pollutions together and mentioned that accurate data of sources is required for reducing industrial pollutions. Industrial activities have main role in producing annual greenhouse gases, about 60%. Among these activities, power stations have significant effect on pollutions, about 21.3% (Valipour et al., 2013b). Some equipment like gas turbines, emit combustion gases to the atmosphere, but there are some useful methods for optimization and environmental pollutants reduction, such as: Output heat recovery from flue gases, Ethylene glycol discharge regulating in the liquid gas dehydration unit and use of chillers to reduce inlet air temperature to turbine and cooling the buildings (Valipour et al., 2013b). In this paper, using mechanical chiller for compressor inlet air temperature reduction, cause air mass flow rate entering the combustion chamber increase. So, as thermal efficiency of gas unit increases (about 1.138% in the warmest month), the more perfect combustion leads to less greenhouse gases emission. 7. Conclusions In order to halter the wasted energy of natural gas in pressure reduction station, a pressure reduction valve substitutes with a turbo expander in the proposed model. In this paper the turbo expander power is used for a mechanical chiller coupled with a heat exchanger to reduce the temperature of compressor inlet air. According to thermal and exergetic analyses of proposed model, the following conclusions can be drawn: 1. By the use of cooling system, air mass flow rate entering the compressor will increase. With assumption of constant amount of natural gas mass flow rate, flue gas mass flow rate entering the gas turbine will raise too. 2. With reducing the compressor required power and increasing the gas turbine output power in proposed model, the plant net output power will improve. The rate of increment is about 1.138% in July which is the warmest month. 3. Supposing constant value of input heat to the combustion chamber, because of increasing the net output power, the thermal efficiency improvement, about 1.138% in July, will become manifest in the proposed model. 4. Improvement of exergetic efficiency and reduction of exergy destruction in total plant by the use of cooling system, are the other advantages of proposed model obtained from exergy analysis. Considering the performance of gas turbine power plant, the model with cooling system has better performance and less environmental pollutions, compared with conventional system. Nomenclatures
Symbol Cp Cv COP
Specific heat at constant pressure (kj/kg K) Specific heat at constant volume (kj/kg K) Coefficient of performance
E_ H LHV M m_ N P Q_ Ru R S S_ T V_ _ W x
r g x hth hc hH.E hex
Rate of exergy flow (kW) Specific enthalpy (kj/kg) lower heating value (kj/kg) Molar mass (kg/mol) Mass flow rate (kg/s) Stoichiometric coefficient of the chemical equation (mol) Pressure (MPa or kPa or bar) Heat transfer rate (kW) Universal gas constant(J/mol K) Compression ratio Specific entropy (kj/kg K) Entropy flow rate (kW/ K) Temperature (K or C) Volume flow rate (m3/s) Actual work rate (kW or MW) Molar fraction Density (kg/m3) Specific heat ratio Chemical exergy/energy ratio Thermal efficiency of cycle (%) Isentropic efficiency of compressor (%) Isentropic efficiency of heat exchanger (%) Exergy efficiency (%)
Subscripts 0 Reference ambient condition 1 Inlet condition of compressor 2 Outlet condition of compressor 3 Inlet gas turbine condition 4 Outlet gas turbine condition 5 Inlet condition of pressure reduction station 6 Inlet condition of pressure reduction valve and expansion turbine 7 Outlet condition of pressure reduction valve and expansion turbine amb Ambient b Basic model c Compressor cc Combustion chamber cv Control volume D Destruction Ext Expansion turbine g Flue gas H.E Heat exchanger i Number of component in Inlet condition L Cooling NG Natural gas out Outlet condition P Proposed model ref Standard state station Pressure reduction station T Gas turbine Th Thermal Superscripts ch Chemical T Thermal M Mechanical W Work or electricity References Abam, F.I., Ugot, I.U., Igbong, D.I., 2012. Performance analysis and components
A. Noroozian, M. Bidi / Journal of Natural Gas Science and Engineering 28 (2016) 95e105 irreversiblities of a (25 MW) gas turbine power plant modeled with a spray cooler. Am. J. Eng. Appl. Sci. 5, 35e41. Ahmadi, P., Dincer, I., Rosen, M.A., 2011. Exergy, exergoeconomic and environmental analyses and evolutionary algorithm based multi-objective optimization of combined cycle power plants. Energy 36 (10), 5886e5898. Al-Doori, W.H.A.R., 2012. Exergy analysis of a gas turbine performance with effect cycle temperatures: IJRRAS. Ameri, M., Hejazi, S.H., 2004. The study of capacity enhancement of the chabahar gas turbine installation using an absorption chiller. Appl. Therm. Eng. 24 (1), 59e68. http://dx.doi.org/10.1016/S1359-4311(03)00239-4. Andrei, I., Valentin, T., Cristina, T., Niculae, T., 2014. Recovery of wasted mechanical energy from the reduction of natural gas pressure. Procedia Eng. 69, 986e990. Badran, O.O., 1999. Gas-turbine performance improvements. Appl. Energy 64 (1e4), 263e273. http://dx.doi.org/10.1016/S0306-2619(99)00088-4. nez-Espadafor, F., Sa nchez, D., S Chacartegui, R., Jime anchez, T., 2008. Analysis of combustion turbine inlet air cooling systems applied to an operating cogeneration power plant. Energy Convers. Manag. 49 (8), 2130e2141. http://dx.doi. org/10.1016/j.enconman.2008.02.023. Chaker, M., Meher-Homji, C.B., Mee, T., Nicholson, A., 2003. Inlet fogging of gas turbine engines detailed climatic analysis of gas turbine evaporation cooling potential in the USA. J. Eng. Gas Turbines Power 125 (1), 300e309. De Lucia, M., Lanfranchi, C., Boggio, V., 1995. Benefits of compressor inlet air cooling for gas turbine cogeneration plants. In: Paper Presented at the ASME 1995 International Gas Turbine and Aeroengine Congress and Exposition. De Sa, A., Al Zubaidy, S., 2011. Gas turbine performance at varying ambient temperature. Appl. Therm. Eng. 31 (14), 2735e2739. Ebadi, M.J., Gorji-Bandpy, M., 2005. Exergetic analysis of gas turbine plants. Int. J. Exergy 2 (1), 31e39. Ehyaei, M., Mozafari, A., Alibiglou, M., 2011. Exergy, economic & environmental (3E) analysis of inlet fogging for gas turbine power plant. Energy 36 (12), 6851e6861. Farzaneh-Gord, M., Deymi-Dashtebayaz, M., 2009. A new approach for enhancing performance of a gas turbine (case study: Khangiran refinery). Appl. Energy 86 (12), 2750e2759. Farzaneh-Gord, M., Deymi-Dashtebayaz, M., 2011. Effect of various inlet air cooling methods on gas turbine performance. Energy 36 (2), 1196e1205. Farzaneh-Gord, M., Izadi, S., Deymi-Dashtebayaz, M., Pishbin, S.I., Sheikhani, H., 2015. Optimizing natural gas reciprocating expansion engines for town border pressure reduction stations based on AGA8 equation of state. J. Nat. Gas Sci. Eng. 26, 6e17. GE Gas turbines. (2006, Dec). Retrieved from: https://www.ge.com/oilandgas.
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Hosseini, R., Beshkani, A., Soltani, M., 2007. Performance improvement of gas turbines of Fars (Iran) combined cycle power plant by intake air cooling using a media evaporative cooler. Energy Convers. Manag. 48 (4), 1055e1064. Ibrahim, T.K., Rahman, M., Abdalla, A.N., 2011. Improvement of gas turbine performance based on inlet air cooling systems: a technical review. Int. J. Phys. Sci. 6 (4), 620e627. Jarze˛ bowski, S., Pyzik, E., Miller, A., 2012. Impact of inlet air cooling on gas turbine performance. J. Power Technol. 92 (4), 249e257. Kakaras, E., Doukelis, A., Prelipceanu, A., Karellas, S., 2006. Inlet air cooling methods for gas turbine based power plants. J. Eng. Gas Turbines Power 128 (2), 312e317. Mohanty, B., Paloso, G., 1995. Enhancing gas turbine performance by intake air cooling using an absorption chiller. Heat Recovery Syst. CHP 15 (1), 41e50. Morini, M., Pinelli, M., Spina, P.R., Vaccari, A., Venturini, M., 2015. Feasibility analysis of gas turbine inlet air cooling by means of liquid nitrogen evaporation for IGCC power augmentation. Appl. Therm. Eng. 80, 168e177. Oyedepo, S.O., Kilanko, O., 2012. Thermodynamic analysis of a gas turbine power plant modelled with an evaporative cooler. Int. J. Thermodyn. 17, 14e20. Po zivil, J., 2004. Use of expansion turbines in natural gas pressure reduction stations. Acta Montan. Slov. 3 (9), 258e260. Shi, X., Agnew, B., Che, D., Gao, J., 2010. Performance enhancement of conventional combined cycle power plant by inlet air cooling, inter-cooling and LNG cold energy utilization. Appl. Therm. Eng. 30 (14), 2003e2010. Tobi, I. Al, 2009. Performance enhancement of gas turbines by inlet air cooling. In: International Conference on Communication, Computer and Power (ICCCP'09) Muscat, February 15e18, pp. 165e170. Valipour, M., Mousavi, S.M., Valipour, R., Rezaei, E., 2012. Air, water, and soil pollution study in industrial units using environmental flow diagram. J. Basic Appl. Sci. Res. 2 (12), 12365e12372. Valipour, M., Mousavi, S.M., Valipour, R., Rezaei, E., 2013. A new approach for environmental crises and its solutions by computer modeling. In: Paper Presented at the 1st International Conference on Environmental Crises and its Solutions, Kish Island, Iran. Valipour, M., Mousavi, S.M., Valipour, R., Rezaei, E., 2013b. Deal with environmental challenges in civil and energy engineering projects using a new technology. J. Civil Environ. Eng. 3 (1), 127e132. Zaki, G.M., Jassim, R.K., Alhazmy, M.M., 2011. Energy, exergy and thermoeconomics analysis of water chiller cooler for gas turbines intake air cooling. Smart Grid Renew. Energy 2 (03), 190. Zeitoun, O., Ali, M., Al-Ansary, H., Nuhait, A., 2014. Ceramic tubes membrane technology as a new humidification technique for gas turbine inlet air cooling. Int. J. Therm. Sci. 80, 1e10.
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Journal of Natural Gas Science and Engineering journal homepage: www.elsevier.com/locate/jngse
An applicable method for gas turbine efficiency improvement. Case study: Montazar Ghaem power plant, Iran Afsaneh Noroozian, Mokhtar Bidi* Faculty of Mechanical and Energy Engineering, Shahid Beheshti University, A.C., Tehran, Iran
a r t i c l e i n f o
a b s t r a c t
Article history: Received 5 September 2015 Received in revised form 19 October 2015 Accepted 16 November 2015 Available online 1 December 2015
Compressor inlet air cooling is one of the most well-known methods to improve gas turbine power plant performance. In this study, pressure reduction valve in natural gas pressure reduction station is replaced by a turbo-expander to use the potential energy of the compressed gas. The turbo-expander shaft is connected to a mechanical chiller to produce refrigeration; the produced refrigeration is used to decrease the compressor inlet air temperature. Moreover, thermodynamic and exergetic analyses are carried out and the effect of compressor air cooling on the performance of the plant is studied. To do so, Montazer Ghaem power plant is considered as the case study. Results showed that using cooling system causes 3.2% temperature drop which leads to 1.138% increment in both thermal efficiency and net output power in the warmest month. Exergetic analysis reveals that using the cooling system leads to a higher exergy efficiency and hence lower exergy destruction. Also combustion chamber with 81.26 MW has the highest amount of exergy destruction which decreases to 77.36 MW after implementation of the cooling system. In January, exergetic efficiency of total plant has 1.86% enhancement and exergy destruction reduces about 3.8 MW. © 2015 Elsevier B.V. All rights reserved.
Keywords: Gas turbine power plant Turbo-expander Compressor inlet air cooling system Efficiency improvement Exergetic analysis
1. Introduction The optimization and reduction of energy consumption has an important role in the contemporary world. Increasing the efficiency of various devices and decreasing their consumed energy is preferred. Gas turbine is a device which has various applications such as producing electricity in power plants, especially those which require less commissioning time. The efficiency and output power of gas turbines are related to many different parameters. The aim of this study is to investigate the influence of ambient temperature on gas turbine plants output power. One Celsius degree temperature increment leads to approximately one percent reduction of the gas turbine rated capacity (Mohanty and Paloso, 1995). There are as many methods to reduce compressor inlet air temperature so as to improve the gas turbine performance. The most important ones are as follows (Kakaras et al., 2006; Hosseini et al., 2007; Chacartegui et al., 2008; Ehyaei et al., 2011):
* Corresponding author. E-mail address: [email protected] (M. Bidi). http://dx.doi.org/10.1016/j.jngse.2015.11.032 1875-5100/© 2015 Elsevier B.V. All rights reserved.
a) Mechanical chiller or absorption chiller: mechanical chiller cools the inlet air so its temperature drops to below the wet bulb temperature. Mechanical chillers are driven by electric motors, steam turbines or engines. These type of chillers use a classical mechanical compression cycle. The inlet air is passed across the cooling coils, through which intermediate refrigerant fluid is circulated. Absorption chillers require steam or hot water as the basic source of energy and they use absorption refrigeration cycle with the working fluids of water/ammonia or lithium bromide/water. Absorption chillers also require less electric energy than mechanical chillers. The main difference of these two types of chillers is in the way that the refrigerant is changed into liquid. This is the reason why absorption chillers do not have any moving parts. b) Chilled water storage or ice harvesting: chilled water storage and ice storage, store sensible and latent heat energy respectively. Ice storage is mostly profitable for small storage capacity. c) Media evaporative cooling and Fogging: the evaporative cooler is a wetted rigid media. As the water distributes from the top, hot air passes through the media and cause water evaporation. So, the air is cooled during the process and then enters the compressor. Fogging is another evaporative cooling technology in which the fine droplets of water are added to the inlet air by
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means of high pressure nozzles. Water consumption in fogging system is more than that of evaporative cooling, but fogging system has higher effectiveness. In a method suggested by Mohanty and Paloso (1995), the heat of flue gas was directly used by an absorption chiller to provide cooling system. Therefore the gas turbine output power increases, the waste heat is recovered and the environmental impacts are decreased. They also concluded that 11% increase in plant's energy output could be achieved. M. De Lucia et al. (1995) discussed differences between absorption and evaporative cooling systems. They concluded that, although evaporative cooling systems have a lower cost, but absorption ones are much more effective especially for base load production plants. Combination of these two cooling systems which has the highest benefit was likewise. Badran (1999) worked on the influence of parameters which could affect the gas turbine performance, such as compressor inlet temperature and turbine inlet temperature. He emphasized that as the compressor inlet air temperature increases, thermal efficiency decreases and also ambient condition would have less effect on efficiency by choosing the best location for power stations. M. Chaker et al. (2003) have studied available direct evaporative cooling hours of 122 locations in the U.S.A with inlet fogging system for gas turbines. Their focus was mostly on climate conditions. Ameri and Hejazi (2004) suggested an air cooling system that uses a steam absorption chiller in Chabahar (Iran) power plant. They concluded that this system with payback period of 4.2 years, would have led to 11.3% increase in average output power. Kakaras et al. (2006) have compared different air cooling methods and their integration. They mentioned that investment costs of evaporative cooling technology are lower than other methods, while absorption chiller system has much higher capacity for increasing electricity generation. Hosseini et al. (2007) evaluated the influence of media evaporative cooling system on gas turbines performance in a combined cycle power plant. Their results showed that this cooling method causes 19 C reduction in inlet air temperature which leads to 11 MW growth in gas turbine output power with a payback period of about 4 years. Shi et al. (2010) performed a research aimed at improvement of combined cycle power plant. They demonstrated that with integration of inlet air cooling, inter cooling and LNG gasification within a combined cycle power station, there would be an increment in net electrical efficiency and output power. De Sa and Al Zubaidy (2011) focused on the relation between ambient temperature and gas turbine efficiency and output power. They mentioned that with increase of temperature, efficiency and output power are reduced. They also established an empirical relationship between ambient humidity and system performance. In addition to conventional cooling methods, there are modern cooling technologies such as ceramic tubes membrane (Zeitoun et al., 2014). Zeitoum et al. worked on an experimental sample and demonstrated that by having both latent and sensible heat transfer, in this model the air passes over the ceramic tube matrix and the water runs through the ceramic tubes. So the ambient temperature decreases and the relative humidity increases with no erosion in compressor blades. There are other methods for gas turbine performance enhancement. Using liquid nitrogen spray is one way to reduce turbine inlet air temperature in Integrated Gasification Combined Cycles (Morini et al., 2015). Some advantages of this method are that it does not need a large amount of water despite common thermal storage systems and there won't be pressure losses at the compressor inlet. Oyedepo and Kilanko (2012) used evaporative cooler for performance enhancement of a gas turbine power plant. They mentioned that 5 C drop in compressor inlet air temperature, causes 5e10% increment in net output power and 2e5% thermal efficiency enhancement. They also showed the positive effect of
pressure ratio on their results. Ibrahim et al. (2011) showed that if compressor inlet air temperature increase 1 C, the gas turbine power output drop about 1%. They also showed that by increasing ambient air temperature, air mass flow rate reduces and finally cycle efficiency decreases. Gas turbine power plants performance in warm and relatively dry climate is very important, so cooling compressor inlet air in these conditions is the topic of many researches (Tobi, 2009; Abam et al., 2012; Jarze˛ bowski et al., 2012). Zaki et al. (2011) demonstrated that by the use of evaporative cooler in hot humid climate, there would be limitation in gas turbine capacity improvement (not more than 5%e7%). Because compressor inlet air temperature should not be less than wet bulb temperature. Farzaneh-Gord and Deymi-Dashtebayaz (2009, 2011) used a similar method for inlet air cooling system to enhance the output power, but they directly used expansion turbine outlet flow for cooling the compressor inlet air by the use of a heat exchanger. Implementation of their model is not possible for a constructed power plant, because their method causes fundamental changes in fuel supply system. As a result, installing of this model needs a long overhaul of power plant and this is against the power plant owners' expectations. On the other hand, because of long distance between pressure reduction station and gas turbine power plant, there would be a possibility to phase change, so this project is very hard to execute. Moreover there were two fundamental shortcomings in their thermodynamic modeling which leads to less accurate results. The first one is that they did not correctly relate the inlet air cooling to output power, another problem is that they used ideal Brayton assumption in their calculations, therefore there will be miscalculations and it could not be used in real industrial decisions. For example Fig. 1 shows differences between the results of simulation procedure with ideal Brayton cycle in comparison with real data for a 134.140 MW gas turbine model (GE, model MS9001E with 34.6% efficiency at shaft output). There is a 7.92% discrepancy between efficiency of ideal Brayton cycle with the catalogue data, while there is no considerable error in the proposed method. In the present study an applicable method for an installed power plant is provided. In the proposed model, a mechanical chiller is used for reducing compressor inlet air temperature which its required power is obtained from a turbo-expander. The turboexpander is replaced in pressure reduction station with a pressure reduction valve. Using a turbo-expander is an appropriate way to recover pressure energy of natural gas in pressure reduction stations (Farzaneh-Gord et al., 2015). Fuel pressure must be reduced from 5e7 MPa to 1.5e4.0 MPa, in the natural gas pressure reduction station (Po zivil, 2004). Andrei et al. (2014) demonstrated that by the use of turbo-expanders instead of throttle valves or pressure reducing valves in pressure reducing stations of the whole country, the sum of power generation could be significant. In this study a preheater is used before the expansion turbine to prevent from production of liquid or solid phase at the discharging of the station. This article compares basic model with proposed model, from the viewpoints of thermodynamics and exergy. As discussed earlier, in previous studies ideal Brayton cycle was used for simulation which caused miscalculations as mentioned in Fig. 1. So in this paper a real cycle is simulated for more accuracy. The proposed model is an applicable method to use potential energy of natural gas unlike previous studies which were impracticable. It is for the first time that exergetic analysis of each component and whole cycle is done for a real gas cycle with cooling system. Exergy destruction rate and exergy efficiency of different points in the cycle are also calculated. Other power plants can use the methodology of the proposed model to enhance the plant power and efficiency, because this model is applicable and does not involve executive problems unlike previous models. Effect of different amounts of COP on net output
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97
Fig. 1. Comparison between results of ideal Brayton cycle and the real gas cycle for GE MS9001E gas turbine.
power is also represented in the results section, so one can see the effects of different condition of cooling system on the output of gas turbine power plants. An upward trend in thermal efficiency, net output power, exergetic efficiency and exergy destruction by the use of proposed model is also predicted for other power plants. However to obtain accurate and applicable results for other power plants, the similar analyses is required using the design data of the power plants and its environmental conditions.
3. Thermal modeling 3.1. Basic model The compressor inlet temperature is equal to ambient temperature in the basic model which neglects the cooling effect and simulates the cycle under ambient conditions.
T1;air;b ¼ Tamb + h1;b ; s1;b P1;air;b ¼ Pamb
2. System description Two different models including basic model and proposed model are investigated in this paper. The basic model consists of a preheater, pressure reduction valve, combustion chamber, compressor and gas turbine as shown in Fig. 2, While Fig. 3 shows the proposed model which has an expansion turbine instead of a pressure reduction valve and it also consists of some auxiliary equipments like heat exchanger and mechanical chiller. Calculation procedures of these two cases are as follows. The flow chart of the basic and proposed model analyses is shown in Fig. 4.
(1)
In this model, inlet air is considered to be an ideal gas and its density is calculated according to the following equation:
r1;air;b ¼
MP1;air;b Ru T1;air;b
(2)
Where M is the molar mass and Ru is the universal gas constant. Volume flow rate of the compressor can be calculated as follow:
Fig. 2. A schematic diagram of the basic model.
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Fig. 3. A schematic diagram of the proposed model.
m_ air;b V_ c;b ¼ r1;air;b
(3)
Wherem_ air;b , is the air mass flow rate of the basic system. The pressure of the air which leaves the compressor (P2,air,b) is specified by the following equation and the temperature of the outlet air (T2,air,b) is obtained from ideal gas equation, considering the value of 0.85 for the isentropic efficiency of the compressor (hc):
P2;air;b ¼ P1;air;b r + 2 3 g1 ! h2;b ; s2;b g T P 5 þ T1;air;b T2;air;b ¼ 1;air;b 4 2;air;b 1 hc P1;air;b
(4)
r is the compression ratio and g is the ratio of specific heat which is calculated using the following formula:
g ¼ Cp=Cv
(5)
Where Cp and Cv are specific heat at constant pressure and volume, respectively. Both of them are calculated at the temperature of T1,air,b. The compressor inlet work is also estimated from:
_ _ air;b h2;b h1;b W c;b ¼ m
(6)
The next step is to calculate the amount of additional heat to preheat natural gas (process5,NG,b_6,NG,b) and also the temperature of natural gas entering the combustion chamber. By applying the first law of thermodynamics:
P5;NG;b ¼ + T5;NG;b ¼ h5;b Q_ in;NG;b ¼ m_ NG;b h6;b h5;b P6;NG;b ¼ P5;NG;b h6;b T6;NG;b Fig. 4. Flow chart of calculation procedure.
(7)
Where m_ NG;b is the mass flow rate of natural gas which passes through the pressure drop station.
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P7;NG;b ¼ P6;NG;b rstation + T7;NG;b h7;b ¼ h6;b
(8)
rstation is the compression ratio of pressure drop station. Using the analysis of the combustion process between the fuel flow and air flow (7,b and 2,b), characteristics of flue gas are defined in point 3. Enthalpy of flue gas in point 3 is calculated using the analysis of the chemical reaction equation for the combustion:
h3;b ¼
X
xi hi
(9)
Analogous to the mass fraction, the mole fraction of component i (xi), is defined as the ratio of the number of moles of i, (ni), to the total number of moles in the mixture, (n):
n xi ¼ i n
(10)
Where n is the stoichiometric coefficient of the chemical equation for the combustion reaction. The heat delivered by the combustion chamber is determined from using the lower heating value of the natural gas (LHV (kj/kg)).
Q_ in;air;b ¼ m_ NG;b LHV
(11)
By assuming the use of a specific turbine, pressure and temperature of point 4 is obtained.
P4;b ¼ T4;b ¼ h4;b ; s4;b
(12)
(13)
Where m_ g;b is the mass flow rate of flue gas and it is given by:
m_ g;b ¼ m_ air;b þ m_ NG;b
(14)
Finally the thermal efficiency of the gas turbine is calculated as follows:
hth;b ¼
_ _ W T;b W c;b _ Q
(16)
So the expansion turbine power is achieved using the following relation:
_ _ NG;p h6;p h7;p W EXT;p ¼ m
(17)
The turbine shaft is connected to a mechanical chiller. Chiller removes the heat from liquid, this liquid can then be circulated through a heat exchanger to cool the air which enters the compressor. The cooling power produced by chiller is equal to:
_ EXT; p COP Ql;chiller ¼ W
(18)
Where COP is the coefficient of performance and in this project, it is considered to be 2.5. This simplification is because of prevention from modeling a refrigeration cycle which causes complicacy of calculations, so a fixed amount is assumed for COP which is suitable for a refrigeration cycle. The next step is to compute the compressor inlet air temperature. As it was mentioned earlier, temperature and pressure in point 1 is equal to ambient conditions. By applying equations related to the heat exchanger, compressor inlet temperature and pressure can be obtained. So the density of air flow is achieved.
Q_ l;chiller hH;E ¼ m_ air;p Cpair T1;air;p T1;c;p
(19)
Where hH,E is the isentropic efficiency of heat exchanger and is assumed to be 0.9. Cpairis the specific heat capacity of air.
The turbine output power is equal to:
_ _ g;b h3;b h4;b W T;b ¼ m
T7;p ¼ P7;p ¼ h7;p ; s7;p
99
(15)
in;air;b
_ _ In the above equation, the term (W T;b W c;b ) shows the output _ power of the whole cycle (W net ).
3.2. Proposed model In the proposed model, some new equipments were added to the cycle, so previous relations should be replaced by new ones. Note that Cp and Cv in the proposed model are calculated at the temperature of T1,c,p. The effect of relative humidity is neglected, because of its yearly average amount which is less than 40% at Montazar Ghaem power plant location and hence it will not cause in much miscalculations. The cooling system is simplified using a constant value for the COP. Within the heat exchanger, the specific heat of air is supposed to be constant. Other model assumptions are listed in Table 2. These quantities are based on Montazar Ghaem's data. By assuming the use of a specific expansion turbine instead of a pressure reduction valve, pressure and temperature of point 7 is obtained.
T1;c;p ¼ P1;c;p M P1;c;p ¼ r1;c;p ¼ R T u 1;c;p
(20)
In the case of installing inlet air cooling system, the inlet compressor pressure is assumed to be unchanged (same as ambient pressure). According to Eq. (20), with decreasing the compressor inlet temperature, the density of air flow similarly, will reduce. Because the volume flow rate of the compressor in both models is constant, so air mass flow rate of proposed model (m_ air;p ) will increase. Here, flow characteristics in different parts of the proposed system are calculated similar to the basic model. By changing flow conditions in compressor and combustion chamber, Eqs. (6), (11), (13)e(15), are rewritten as follows:
_ c;p ¼ m_ W air;p h2;p h1;p
(21)
Q_ in;air;p ¼ m_ NG;p LHV
(22)
_ _ g;p h3;p h4;p W T;p ¼ m
(23)
Where m_ g;p is the mass flow rate of flue gas and it is given by:
m_ g;p ¼ m_ air;p þ m_ NG;p
hth;p ¼
_ _ W T;p W c;p _ Q in;air;p
(24)
(25)
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3.3. Exergy balance equation
Exergy is the maximum amount of useful work which can be obtained from a machine. Exergy won't be stored during a process, but it destroys because of irreversibility. Each single component of a cycle can be analyzed individually in this method and each component's contribution to exergy's waste in the whole system could be obtained. An exergetic analysis was derived for each component of GE-MS9001 gas turbine power plant. Exergetic efficiency and exergy destruction of cycle with compressor inlet air cooling system and without it were compared. In this part, equations are written for an ideal gas having a constant specific heat. The thermo-mechanical exergy flow can be divided into two branches, thermal and mechanical exergy flows (Al-Doori, 2012), which are related to the temperature and pressure of the stream. Chemical exergy is related to the composition of the fuel stream. The general exergy balance equation is as follows (Ebadi and GorjiBandpy, 2005): ch E_ þ
X
T E_ in
out w
! T E_ out
out
in
X
X
. S_out þ Q_ cv T ref
þ
!
X
M E_ in
X
! M E_ out
out
in
þ Tref
X
S_in
in
(26) In this equation, subscripts in and out respectively show the inlet and outlet of the exergy flow through each component. Subscript ref indicates reference environmental conditions which are:Pref ¼ 100 (kp), Tref ¼ 25 ( C). Ru is universal gas constant and equals to 0.287 (kj/kg.k).Q_ cv shows the heat transfer between component and environment. P P _ The term Tref ð S_in Sout Þ shows exergy destruction rate out in (E_ D ). The thermal and mechanical exergy flow rates can be written as follows (Ebadi and Gorji-Bandpy, 2005):
"
T Tref
M E_ ¼ m_ R Tref ln
P Pref
Tref
T ln Tref
!# (27)
! (28)
For chemical exergy, following approximation is used (Ahmadi et al., 2011): ch E_ ¼ x LHV m_ NG 1000
(29)
=
Where x is chemical exergy/energy ratio and in this study, its value for CH4 is 1.06.
3.3.1. Exergy balance and exergy efficiency equations for components Exergy balance and exergy efficiency equations for each component are as follows: (Al-Doori, 2012): Compressor:
M T T M _c E_ 1 E_ 2 þ E_ 1 E_ 2 þ Tref S_1 S_2 ¼ W
(30)
_c W
(31)
Combustion chamber:
M T ch T T M M E_ þ E_ 2 þ E_ 7 E_ 3 þ E_ 2 þ E_ 7 E_ 3 þ Tref S_2 þ S_7 S_3 . þ Q_ cv T ref
!
¼0 (32)
hex;cc
T M E_ 3 þ E_ 3 ¼ T T M M ch E_ 2 þ E_ 2 þ E_ 7 þ E_ 7 þ E_
(33)
Gas turbine:
T M T M _T E_ 3 E_ 4 þ E_ 3 E_ 4 þ Tref S_3 S_4 ¼ W
hex;c ¼
¼ E_
T E_ ¼ m_ Cp
hex;c ¼
T T M M E_ 2 þ E_ 2 E_ 1 þ E_ 2
T E_ 3
þ
M E_ 3
_ W T T M E_ 4 þ E_ 4
(34)
(35)
The same equations could be used to calculate exergy performance of the system with cooling system. 4. Case study Montazar Ghaem power plant which is located at 7 km Mallard road in Alborz province, consists of different power generator plants as follows: Four steam power generators, six gas turbine generators and three combined cycle generators. This plant has a total capacity of 625.2 MW. Each combined cycle power unit includes two GEMs9001E gas turbines, with nominal capacity of 110 MW. The technical data of these turbines are shown in Table 1. In this study, the validity of proposed model will be checked by Montazar Ghaem's data. Table 2 shows the recorded values in Montazar Ghaem plant. Table 3 shows average monthly temperatures of Alborz province where the power plant is located. 5. Results and discussions For verification of the developed code, the modeling results were compared with the obtained results of Ameri and Hejazi (2004) and a good agreement between these two studies were found. One of Montazar Ghaem's gas units was selected as case study for thermal and exergetic analysis. In this section, the effects of inlet air cooling on gas turbine performance are investigated. As discussed earlier, the proposed model is implemented on the GEMS9001E gas turbine which is installed within an existing power plant in Iran. Figs. 5e10 show the effects of temperature drop on gas cycle characteristic parameters. By the use of a mechanical chiller coupled with a heat exchanger, the temperature of compressor inlet air can be reduced as shown in Fig. 5. According to Eq. (17) and Eq. (18), cooling power depends on the coefficient of performance (COP) and reduction station inlet and outlet conditions. Therefore by assuming a
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Table 1 Feature of gas turbine engine used for simulation (ISO conditions, natural gas) (GE Gas turbines, 2006). Ms9001E
ISO rated power (kW)
Heat rate (kJ/kWh)
Efficiency (%)
Pressure ratio
Exhaust temperature ( C)
Exhaust flow (kg/s)
134,140
10,397
34.6
12.7
540
410.0
Table 2 Montazar Ghaem's data which is taken from its official Website. Description
Recorded value
Unit
Mass flow rate of air, entering the compressor (m_ air ) Mass flow rate of natural gas through of the pressure drop station (m_ NG ) compression pressure ratio (r) Ambient temperature (Tamb) Ambient (compressor inlet) pressure (P1,air) Inlet pressure of the pressure reduction station (P5,NG) Outlet pressure of the pressure reduction station (P7,NG) Inlet temperature of the pressure reduction station (T5,NG) Outlet temperature of the pressure reduction station (T7,NG) Outlet temperature of the gas turbine (T4) Specific heat at constant pressure (Cpair)
350 4.16 10:1 25 0.9 51.7 17.2 60 11 525 1.004
(kg/s) (kg/s) e ( C) (bar) (bar) (bar) ( C) ( C) ( C) (kj/kg K)
The analysis carried out by taking the weather data of Iran Meteorological Organization for Alborz province weather, into account.
Table 3 Average monthly temperature of Alborz (Iran) province ( C) which is taken from I.R. of Iran Meteorological Organization Website. Year
Jan
Feb
Mar
Apr
May
Jun
Jul
Aug
Sep
Oct
Nov
Dec
2014
3.6
3.5
10.4
16
21.8
26.3
28.9
28.7
24.8
15.1
7.7
5.9
Fig. 7. Compressor required power.
Fig. 5. Compressor inlet temperature.
Fig. 8. Gas turbine output power.
Fig. 6. Turbine air mass flow rate.
constant value for the coefficient of performance and reduction station conditions, a constant value of cooling power and a constant temperature drop during a year are obtained in the proposed model. Temperature drop is about 3.2% during a year.
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Fig. 9. Net output power of cycle.
Fig. 10. Gas turbine thermal efficiency.
Fig. 6 shows that inlet air cooling will make the air mass flow rate increase. Supposing that the air pressure is constant in Eq. (20), the air density increases by reduction of air temperature. Therefore, with this supposition that the air volume flow rate (V_ c ) is constant, the air mass flow rate (m_ air ) in the proposed model will increase. The air mass flow rate in warm seasons is less than cold seasons because air density decreases in warm seasons. The percent of increase in the air mass flow rate due to the use of proposed model is about 3.39 in July and 3.52 in February. As it is shown in Figs. 7e8 it's clear that the influence of temperature drop on gas turbine power is more than its influence on compressor required power, but for both of them, changes are not tangible. Change in compressor inlet air temperature, causes alteration in specific enthalpy. As is construed from Eq. (4), any changes in the specific heat ratios and compressor inlet temperature leads to variation in temperature and specific enthalpy of point 2 rather than its value in basic model. Variation of specific heat ratios (g) is due to different quantities of Cp and Cv in the proposed model and the variation of Cp and Cv is related to compressor inlet temperature drop. Finally, according to Eq. (21), in the model with cooling system, enthalpy difference decreases more than the air mass flow rate increment. As a result the required power of compressor decreases a bit. Reduction of compressor required power in warmer months like July (0.18%) is a bit less than cold months (0.26% in February). Since the mass flow rate of flue gas entering the gas turbine is the sum of air mass flow rate and natural gas mass flow rate, with the increase of air mass flow rate and constant amount of natural gas mass flow rate, flue gas mass flow rate will increase. By the use of a special gas turbine, thermodynamic parameters of point 4 remain constant in both models. Specific enthalpy of point 3 is obtained from combustion analysis and is fixed in both models.
According to Eq. (13), the amount of gas turbine output power is multiplication of enthalpy difference of points 3 to 4 and flue gas mass flow rate. In the proposed model, the enthalpy difference is unchanged and as the flue gas mass flow rate increases, the gas turbine output power also increases. Rate of changes in July and February are 0.45% and 0.33% respectively. The net output power and thermal efficiency of gas turbine cycle increase with the decrease of compressor inlet temperature, as depicted in Figs. 9e10 respectively. The net output power is the difference between gas turbine output power and compressor required power. So as discussed earlier, in the proposed model, gas turbine output power will increase till the compressor required power decreases and finally net output power increases subsequently. As shown in Fig. 9. Net output power is increased by 1.138% and 0.98% for July and February respectively. Since there are 6 gas units in this power plant, therefore a total increase in power will be 6484.23 kW in July and 5975.88 kW in February. Significant increment in thermal efficiency of proposed model can be explained by Eq. (25). Thermal efficiency is related to the net output power and the heat delivered by the combustion chamber. The heat delivered by the combustion chamber (Q_ in;air ) is determined from multiplication of the lower heating value of the natural gas (LHV) and its mass flow rate (m_ NG ), so with unchanged amount of these parameters in both models, the amount of delivered heat remains constant. With the increase of net output power in the model with cooling system, thermal efficiency will improve. The amounts of efficiency alterations are as same as net output power alterations, so the percentages of changes are 1.138% and 0.98% in July and February respectively. Fig. 11 shows the net output power quantities in basic and proposed model compare with real measured data of the power plant. The differences between the real data and results of the basic model show the simulation errors. It is clear that the cycle with cooling system proposed more net output power than the others. Deviations from real quantities in warm seasons are more than cold seasons especially in July and August, as depicted in Fig. 11. Effect of different amounts of COP on the net output power is shown in Fig. 12. It is clear that with decrease of COP in proposed model, net output power will be decreased. According to Eq. (18), the cooling power produced by chiller (Q_ l;chiller ) is proportional to the COP, so with lowering the COP, Q_ l;chiller will also decrease and its positive effects on the gas cycle parameters will reduce. Optimizing the cooling system parameters like COP, can be done using optimization methods such as genetic or PSO algorithms with selection of an objective function such as maximization of exergetic efficiency or net output power. To do this an accurate modeling of the cooling system is required and is not the main objective of the present study.
Fig. 11. Net output power of basic and proposed models in comparison with real quantities.
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Table 5 Net exergy flow rates and exergy destruction in the gas turbine power plant at rated conditions (Without cooling system). Component
_ ðMWÞ W
ch E_ ðMWÞ
T E_ ðMWÞ
M E_ ðMWÞ
E_ D ðMWÞ
Compressor Combustion chamber Gas turbine Total cycle
61.604 0 117.940 56.336
0 220.5 0 220.5
28.53 139.3 119.9 47.93
51.35 0.063 52.30 1.013
18.276 80.763 54.26 153.3
Table 6 Property values and thermal, mechanical and chemical exergy flow rates at various state points in the gas turbine cycle at rate conditions (With cooling system).
Fig. 12. Effect of COP on net output power of proposed model.
5.1. Exergetic analysis results Tables 4 and 6, show property values at various points in the cycle, with consideration of cooling system and without it. These parameters were measured in previous section. Tables 5 and 7, show net exergy flow rates and exergy destruction of control volumes that have been assumed around each component, with consideration of cooling system and without it. For exergetic analysis, data of one month, January, is used. The values which are shown in Tables 4 and 6 are obtained using the thermodynamic properties of each point such as pressure, temperature and mass flow rate. The analysis of different exergy streams in a system can be done using the thermophysical figures of JANAF Tables (Ebadi and Gorji-Bandpy, 2005). There's important information in Tables 5 and 7 for the comparison of the two models. The resources which generate the exergy of the cycle are shown using negative values. Fuel also has negative exergy value. Exergy flow rate of products are shown positive (Ebadi and Gorji-Bandpy, 2005). For satisfying the exergy balance of each component and also the whole cycle, the sum of exergy flow rates of recourses, products and destruction should be zero. Fig. 13 shows exergetic efficiency enhancement by the use of cooling system. In January, temperature decreases from 276.75 to 267.4 K and this reduction causes exergy efficiency of total plant increases from 57.37% to 58.44%. It is clear that efficiency increment in combustion chamber is the most, compared with compressor and gas turbine. The exergy efficiency of combustion chamber experienced an increase from 73.07% to 74.33%. Decrease of exergy destruction in total plant by the use of cooling system is shown in Fig. 14. Exergy destruction in combustion chamber has the highest amount in comparison with compressor and gas turbine. This is because combustion chamber has greater irreversibility than other components. According to temperature reduction, exergy destruction of total plant reduced from 153.3 to 149.5 MW. Also, as observed in Fig. 14, the decline in the exergy destruction of the combustion chamber is remarkable.
State point
m_ ðkg=sÞ
T (k)
P (bar)
ch E_ ðMWÞ
T E_ ðMWÞ
M E_ ðMWÞ
1c 2 3 4 7
215.290 215.290 219.450 219.450 4.160
267.4 645.6 1405.6 798.15 284.15
1 17.9 17.9 1 17.2
0 0 0 0 220.5
0.369 26.77 169.9 49.75 0.00142
0 53.15 54.17 0 1.013
Table 7 Net exergy flow rates and exergy destruction in the gas turbine power plant at rated conditions (With cooling system). Component
_ ðMWÞ W
ch E_ ðMWÞ
T E_ ðMWÞ
M E_ ðMWÞ
E_ D ðMWÞ
Compressor Combustion chamber Gas turbine Total cycle
61.444 0 118.336 56.892
0 220.5 0 220.5
26.4 143.1 120.1 49.38
53.15 0.007 54.17 1.013
18.106 77.4 54 149.506
Fig. 13. Exergetic efficiency of components and total plant.
Table 4 Property values and thermal, mechanical and chemical exergy flow rates at various state points in the gas turbine cycle at rate conditions (Without cooling system). State point
m_ ðkg=sÞ
T (k)
P (bar)
ch E_ ðMWÞ
T E_ ðMWÞ
M E_ ðMWÞ
1 2 3 4 7
208 208 212.160 212.160 4.16
276.75 667.9 1422.869 798.15 284.15
1 17.9 17.9 1 17.2
0 0 0 0 220.5
0.1689 28.7 168 48.1 0.00142
0 51.35 52.30 0 1.013
Fig. 14. Exergy destruction in components and total system.
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6. Environmental impacts Nowadays, environmental impacts are matter of concern to industry owners. Many studies have been done about ways to reduce industrial pollutions using classical methods (Valipour et al., 2013a). Valipour et al. (2012) designed an environmental flow diagram (EFD) for determining environmental impacts of industries. They studied on air, water and soil pollutions together and mentioned that accurate data of sources is required for reducing industrial pollutions. Industrial activities have main role in producing annual greenhouse gases, about 60%. Among these activities, power stations have significant effect on pollutions, about 21.3% (Valipour et al., 2013b). Some equipment like gas turbines, emit combustion gases to the atmosphere, but there are some useful methods for optimization and environmental pollutants reduction, such as: Output heat recovery from flue gases, Ethylene glycol discharge regulating in the liquid gas dehydration unit and use of chillers to reduce inlet air temperature to turbine and cooling the buildings (Valipour et al., 2013b). In this paper, using mechanical chiller for compressor inlet air temperature reduction, cause air mass flow rate entering the combustion chamber increase. So, as thermal efficiency of gas unit increases (about 1.138% in the warmest month), the more perfect combustion leads to less greenhouse gases emission. 7. Conclusions In order to halter the wasted energy of natural gas in pressure reduction station, a pressure reduction valve substitutes with a turbo expander in the proposed model. In this paper the turbo expander power is used for a mechanical chiller coupled with a heat exchanger to reduce the temperature of compressor inlet air. According to thermal and exergetic analyses of proposed model, the following conclusions can be drawn: 1. By the use of cooling system, air mass flow rate entering the compressor will increase. With assumption of constant amount of natural gas mass flow rate, flue gas mass flow rate entering the gas turbine will raise too. 2. With reducing the compressor required power and increasing the gas turbine output power in proposed model, the plant net output power will improve. The rate of increment is about 1.138% in July which is the warmest month. 3. Supposing constant value of input heat to the combustion chamber, because of increasing the net output power, the thermal efficiency improvement, about 1.138% in July, will become manifest in the proposed model. 4. Improvement of exergetic efficiency and reduction of exergy destruction in total plant by the use of cooling system, are the other advantages of proposed model obtained from exergy analysis. Considering the performance of gas turbine power plant, the model with cooling system has better performance and less environmental pollutions, compared with conventional system. Nomenclatures
Symbol Cp Cv COP
Specific heat at constant pressure (kj/kg K) Specific heat at constant volume (kj/kg K) Coefficient of performance
E_ H LHV M m_ N P Q_ Ru R S S_ T V_ _ W x
r g x hth hc hH.E hex
Rate of exergy flow (kW) Specific enthalpy (kj/kg) lower heating value (kj/kg) Molar mass (kg/mol) Mass flow rate (kg/s) Stoichiometric coefficient of the chemical equation (mol) Pressure (MPa or kPa or bar) Heat transfer rate (kW) Universal gas constant(J/mol K) Compression ratio Specific entropy (kj/kg K) Entropy flow rate (kW/ K) Temperature (K or C) Volume flow rate (m3/s) Actual work rate (kW or MW) Molar fraction Density (kg/m3) Specific heat ratio Chemical exergy/energy ratio Thermal efficiency of cycle (%) Isentropic efficiency of compressor (%) Isentropic efficiency of heat exchanger (%) Exergy efficiency (%)
Subscripts 0 Reference ambient condition 1 Inlet condition of compressor 2 Outlet condition of compressor 3 Inlet gas turbine condition 4 Outlet gas turbine condition 5 Inlet condition of pressure reduction station 6 Inlet condition of pressure reduction valve and expansion turbine 7 Outlet condition of pressure reduction valve and expansion turbine amb Ambient b Basic model c Compressor cc Combustion chamber cv Control volume D Destruction Ext Expansion turbine g Flue gas H.E Heat exchanger i Number of component in Inlet condition L Cooling NG Natural gas out Outlet condition P Proposed model ref Standard state station Pressure reduction station T Gas turbine Th Thermal Superscripts ch Chemical T Thermal M Mechanical W Work or electricity References Abam, F.I., Ugot, I.U., Igbong, D.I., 2012. Performance analysis and components
A. Noroozian, M. Bidi / Journal of Natural Gas Science and Engineering 28 (2016) 95e105 irreversiblities of a (25 MW) gas turbine power plant modeled with a spray cooler. Am. J. Eng. Appl. Sci. 5, 35e41. Ahmadi, P., Dincer, I., Rosen, M.A., 2011. Exergy, exergoeconomic and environmental analyses and evolutionary algorithm based multi-objective optimization of combined cycle power plants. Energy 36 (10), 5886e5898. Al-Doori, W.H.A.R., 2012. Exergy analysis of a gas turbine performance with effect cycle temperatures: IJRRAS. Ameri, M., Hejazi, S.H., 2004. The study of capacity enhancement of the chabahar gas turbine installation using an absorption chiller. Appl. Therm. Eng. 24 (1), 59e68. http://dx.doi.org/10.1016/S1359-4311(03)00239-4. Andrei, I., Valentin, T., Cristina, T., Niculae, T., 2014. Recovery of wasted mechanical energy from the reduction of natural gas pressure. Procedia Eng. 69, 986e990. Badran, O.O., 1999. Gas-turbine performance improvements. Appl. Energy 64 (1e4), 263e273. http://dx.doi.org/10.1016/S0306-2619(99)00088-4. nez-Espadafor, F., Sa nchez, D., S Chacartegui, R., Jime anchez, T., 2008. Analysis of combustion turbine inlet air cooling systems applied to an operating cogeneration power plant. Energy Convers. Manag. 49 (8), 2130e2141. http://dx.doi. org/10.1016/j.enconman.2008.02.023. Chaker, M., Meher-Homji, C.B., Mee, T., Nicholson, A., 2003. Inlet fogging of gas turbine engines detailed climatic analysis of gas turbine evaporation cooling potential in the USA. J. Eng. Gas Turbines Power 125 (1), 300e309. De Lucia, M., Lanfranchi, C., Boggio, V., 1995. Benefits of compressor inlet air cooling for gas turbine cogeneration plants. In: Paper Presented at the ASME 1995 International Gas Turbine and Aeroengine Congress and Exposition. De Sa, A., Al Zubaidy, S., 2011. Gas turbine performance at varying ambient temperature. Appl. Therm. Eng. 31 (14), 2735e2739. Ebadi, M.J., Gorji-Bandpy, M., 2005. Exergetic analysis of gas turbine plants. Int. J. Exergy 2 (1), 31e39. Ehyaei, M., Mozafari, A., Alibiglou, M., 2011. Exergy, economic & environmental (3E) analysis of inlet fogging for gas turbine power plant. Energy 36 (12), 6851e6861. Farzaneh-Gord, M., Deymi-Dashtebayaz, M., 2009. A new approach for enhancing performance of a gas turbine (case study: Khangiran refinery). Appl. Energy 86 (12), 2750e2759. Farzaneh-Gord, M., Deymi-Dashtebayaz, M., 2011. Effect of various inlet air cooling methods on gas turbine performance. Energy 36 (2), 1196e1205. Farzaneh-Gord, M., Izadi, S., Deymi-Dashtebayaz, M., Pishbin, S.I., Sheikhani, H., 2015. Optimizing natural gas reciprocating expansion engines for town border pressure reduction stations based on AGA8 equation of state. J. Nat. Gas Sci. Eng. 26, 6e17. GE Gas turbines. (2006, Dec). Retrieved from: https://www.ge.com/oilandgas.
105
Hosseini, R., Beshkani, A., Soltani, M., 2007. Performance improvement of gas turbines of Fars (Iran) combined cycle power plant by intake air cooling using a media evaporative cooler. Energy Convers. Manag. 48 (4), 1055e1064. Ibrahim, T.K., Rahman, M., Abdalla, A.N., 2011. Improvement of gas turbine performance based on inlet air cooling systems: a technical review. Int. J. Phys. Sci. 6 (4), 620e627. Jarze˛ bowski, S., Pyzik, E., Miller, A., 2012. Impact of inlet air cooling on gas turbine performance. J. Power Technol. 92 (4), 249e257. Kakaras, E., Doukelis, A., Prelipceanu, A., Karellas, S., 2006. Inlet air cooling methods for gas turbine based power plants. J. Eng. Gas Turbines Power 128 (2), 312e317. Mohanty, B., Paloso, G., 1995. Enhancing gas turbine performance by intake air cooling using an absorption chiller. Heat Recovery Syst. CHP 15 (1), 41e50. Morini, M., Pinelli, M., Spina, P.R., Vaccari, A., Venturini, M., 2015. Feasibility analysis of gas turbine inlet air cooling by means of liquid nitrogen evaporation for IGCC power augmentation. Appl. Therm. Eng. 80, 168e177. Oyedepo, S.O., Kilanko, O., 2012. Thermodynamic analysis of a gas turbine power plant modelled with an evaporative cooler. Int. J. Thermodyn. 17, 14e20. Po zivil, J., 2004. Use of expansion turbines in natural gas pressure reduction stations. Acta Montan. Slov. 3 (9), 258e260. Shi, X., Agnew, B., Che, D., Gao, J., 2010. Performance enhancement of conventional combined cycle power plant by inlet air cooling, inter-cooling and LNG cold energy utilization. Appl. Therm. Eng. 30 (14), 2003e2010. Tobi, I. Al, 2009. Performance enhancement of gas turbines by inlet air cooling. In: International Conference on Communication, Computer and Power (ICCCP'09) Muscat, February 15e18, pp. 165e170. Valipour, M., Mousavi, S.M., Valipour, R., Rezaei, E., 2012. Air, water, and soil pollution study in industrial units using environmental flow diagram. J. Basic Appl. Sci. Res. 2 (12), 12365e12372. Valipour, M., Mousavi, S.M., Valipour, R., Rezaei, E., 2013. A new approach for environmental crises and its solutions by computer modeling. In: Paper Presented at the 1st International Conference on Environmental Crises and its Solutions, Kish Island, Iran. Valipour, M., Mousavi, S.M., Valipour, R., Rezaei, E., 2013b. Deal with environmental challenges in civil and energy engineering projects using a new technology. J. Civil Environ. Eng. 3 (1), 127e132. Zaki, G.M., Jassim, R.K., Alhazmy, M.M., 2011. Energy, exergy and thermoeconomics analysis of water chiller cooler for gas turbines intake air cooling. Smart Grid Renew. Energy 2 (03), 190. Zeitoun, O., Ali, M., Al-Ansary, H., Nuhait, A., 2014. Ceramic tubes membrane technology as a new humidification technique for gas turbine inlet air cooling. Int. J. Therm. Sci. 80, 1e10.