Thermodynamic assessment of a wind turbine based combined cycle

Energy 44 (2012) 321e328

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Thermodynamic assessment of a wind turbine based combined cycle M. Rabbani, I. Dincer, G.F. Naterer* Faculty of Engineering and Applied Science, University of Ontario Institute of Technology (UOIT), 2000 Simcoe Street North, Oshawa, ON, Canada L1H 7K4

a r t i c l e i n f o

a b s t r a c t

Article history: Received 11 January 2012 Received in revised form 9 May 2012 Accepted 9 June 2012 Available online 11 July 2012

Combined cycles use the exhaust gases released from a Gas Turbine (GT). Approximately 30e40% of the turbine shaft work is typically used to drive the Compressor. The present study analyzes a system that couples a Wind Turbine (WT) with a combined cycle. It demonstrates how a WT can be used to supply power to the Compressor in the GT cycle and pump fluid through a reheat Rankine cycle, in order to increase the overall power output. Three different configurations are discussed, namely high penetration, low penetration and wind power addition. In the case of a low electricity demand and high penetration configuration, extra wind power is used to compress air which can then be used in the low penetration configuration. During a high load demand, all the wind power is used to drive the pump and compressor and if required additional compressed air is supplied by a storage unit. The analysis shows that increasing the combustion temperature reduces the critical velocity and mass flow rate. Increases in wind speed reduce both energy and exergy efficiency of the overall system. Ó 2012 Elsevier Ltd. All rights reserved.

Keywords: Combined wind energy system Thermodynamics Exergy Efficiency

1. Introduction Combined power plants are used to increase the efficiency of a system. They use the exhaust gases released from Gas Turbines (GT). These gases mostly consist of CO2, H2O, N2, and O2 at very high temperatures (around 500e600
C) and low pressures. These temperatures are often high enough to drive a Rankine cycle. Combined power plants also have higher efficiencies when compared to gas and steam power plants. Compressors and pumps are usually driven by turbine power output, which reduces the net work output of the system. Usually 30e40% of the GT work is used to drive the compressor. Exergy analysis is a valuable tool in both the thermodynamic analysis and design of power plants. Exergy can be defined as the maximum obtainable work from a system before it reaches equilibrium with the surroundings. In order to maximize the power plant output, it is important to determine the degree of irreversibility within plant components. Exergy analysis is a valuable tool to analyze inefficiencies within plant components [1e4]. Maeero et al. [5] have analyzed and optimized the exergy processes of combined triple power plants. The triple analysis includes analysis of the Brayton cycle (gas-based) and two Rankine cycles (steam and ammonia-based). The results showed that the maximum exergy

* Corresponding author. E-mail addresses: [email protected] (M. Rabbani), Ibrahim.Dincer@ uoit.ca (I. Dincer), [email protected] (G.F. Naterer). 0360-5442/$ e see front matter Ó 2012 Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.energy.2012.06.027

destruction occurs in the heat recovery heat exchanger. Their study demonstrated that the use of feedwater heaters results in increasing the efficiency. Also, with increases in ambient temperature, the exergy efficiency decreases and with increases in pressure, the ratio of exergy efficiency increases up to a certain level and then decreases. Zheng et al. [6] have investigated a combined power and refrigeration cycle. Their simulation results showed that the proposed cycle has potential to produce a refrigeration effect and most of the exergy losses occur in the ejector. In another study, Zhang et al. [7] studied a combined Brayton and inverse Brayton cycle. Results demonstrated that optimal efficiency will be achieved by controlling the mass flow rate of air, water and fuel. Spelling et al. [8] presented the integration of a solar unit and combined cycle. A thermo-economic analysis showed that efficiencies of 18e24% can be achieved. Ahmadi et al. [9] performed energy, exergy, exergoeconomic and environmental analyses of a combined power plant. The results indicated that with increases in the pressure ratio, the steam cycle power decreased. This ultimately increased the overall efficiency of the system and CO2 emissions decreased. Also, CO2 emissions can be reduced by selecting the best component with a low fuel injection rate. Dincer and Al-Muslim [10] analyzed a Rankine cycle with reheat and found the energy and exergy efficiencies by varying the operating conditions. Rosen and Dincer [11] performed energy and exergy analyses of various industrial processes. Cihan et al. [12] performed an exergy analysis of a power plant in Turkey using a parametric study of different components. Results demonstrated that the CC, GT and heat recovery heat exchanger are the plant

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Nomenclature A ex E_ xd h _ m P Q_ R s S_ g T V _ W y ye

Area of wind turbine (m2) Specific enthalpy, kJ/kg Exergy destruction rate, kW Specific enthalpy, kJ/kg Mass flow rate, kg/s Pressure, kPa Heat flow rate, kW Universal gas constant, kJ K1 mol1 Specific entropy, kJ/kg-K Entropy generation rate, kW/K Temperature, K Velocity (m/s) Work rate, kW Mole fraction Mole fraction in the environment

Greek letters h Efficiency 4 Betz limit r Density of air

components with the highest exergy destruction. Ameri et al. [13] performed energy, exergy and exergoeconomics analyses of a power plant in Iran. The results showed that most of the exergy destruction occurs in the condenser (i.e. 307 MW) while the boiler contributed up to 68 MW. Ameri et al. [14] showed that exergy destruction in the heat recovery heat exchanger can be reduced if a duct burner is incorporated. Ibrahim et al. [15,16] studied the integration of Wind Turbines (WT) with a diesel cycle for remote areas in Canada. Several different cases were examined and it was concluded that integration of WTs with the diesel cycle reduces the environmental impact and improves the overall efficiency of the system. With high penetration, a compressed air storage option was used, which can also be used during low penetration periods. In a separate study, the authors also performed optimization of diesel engines integrated with wind power [17]. Crawford [18] has studied life cycle energy and greenhouse emission analysis of WTs by considering their size and energy yield. After analyzing two different size turbines, the results show that there is no considerable difference in the energy yield of small and big turbines, however using a large WT reduces the footprint area per unit of rated power. Pope et al. [19] presented energy and exergy analyses of vertical and horizontal WTs. The results showed 50e53% differences in the energy and exergy efficiencies for horizontal WT and 44e55% differences in the case of vertical WTs. Also better site selection and turbine design through an exergy method can increase the production capacity of the WT and decrease the economic cost. Hall et al. [20] presented a variable ratio gear box in a WT to increase the efficiency of the system. From the results, a variable ratio gear box can benefit all types of WTs regardless of their type and also improve the efficiency at relatively low cost. The objective of the present study is to thermodynamically analyze a combined system (i.e. Brayton plus reheat Rankine) coupled with a WT. A WT is used to supply power to the compressor in the Brayton cycle and pump fluid in the Rankine cycle. Exergy analysis is performed on the overall system and for individual system components. Different plant parameters will be varied, after which results will be presented and discussed.

Acronyms HHV Higher heating value LHV Lower heating value Subscripts c critical CD Condenser Cog Cogeneration Comp Compressor CC Combustion chamber en Energy ex Exergy f Fuel GT Gas turbine HPT High pressure turbine HX Heat exchanger LPT Low pressure turbine m Motor 0 Reference state P Pump s space tf Thermophysical WT Wind turbine

2. System description A 100 MW combined power plant will be analyzed in this section. This system consists of a combined Brayton cycle and reheat Rankine cycle. The Brayton cycle consists of an air compressor, CC and GT. The exhaust of the Brayton cycle at high temperature and low pressure is used to drive a reheat Rankine cycle. The reheat Rankine cycle consists of a pump, high and low pressure steam turbines, heat recovery heat exchanger and condenser. The required input power for the compressor in the Brayton cycle and the pump in the Rankine cycle is provided by the WT. The compressor is connected with the Brayton cycle and the pump is connected with the reheat Rankine cycle. Heat rejected from the condenser is used to heat water, which could potentially be used in winters for heating purposes. The system is dependent on the amount of wind power available to drive the compressor and the pump. Three different cases are studied; namely high penetration, low penetration, and wind power addition to the combined power plant. Fig. 1 shows a schematic of the system analyzed in the present study. Two flow valves, namely V1 and V2, are used in the system and both operate differently depending upon the mode of operation and power controller. Fig. 2 shows the sequence of the power controller. 2.1. High penetration system In the case of the high penetration system, it is assumed that the WT produces more electrical power output than required by the compressor and pump. So in this case, the wind speed is above the _ 5 ¼ 0Þ while V1 is closed critical value. A flow valve V2 is opened ðm _ 18 ¼ 0Þ. This extra energy is used to store compressed air which ðm can later be supplied to the CC in the case of low wind power. Fig. 3a and b shows the energetic and exegetic high penetration configuration. 2.2. Low penetration system In the case of the low penetration system, WT does not supply enough power to the compressor and pump, so in order to maintain

M. Rabbani et al. / Energy 44 (2012) 321e328

323

Fig. 1. System schematic.

the air to fuel ratio, compressed air from a storage tank is used. _ 18 > 0Þ. _ 5 ¼ 0Þ while V1 is opened ðm Valve V2 is closed ðm Compressed air from the storage tank is preheated in HX2 by the exhaust going out of the HRSG and entering into the CC. Fig. 4a and b show the energetic and exegetic low penetration configuration. 2.3. Wind power addition In the case of the wind power addition, the system is analyzed by adding the wind power to the output while the pump and the compressor are powered by the combined cycle. Fig. 5a and b shows the energetic and exegetic wind power addition configuration. 3. Thermodynamic analysis of the system The critical velocity is the minimum wind speed which produces enough wind power to drive the pump and compressor without

any storage and also without any input from the storage unit. In other words, it is the degree of wind speed above which the plant is working in the high penetration mode and below which it is _ 18 ¼ 0, the _5 ¼ m working in the low penetration mode. When m wind speed at that point is known as the critical velocity. It is dependent on the load demand. The critical mass flow rate is the minimum mass flow rate of the air input to the CC in order to meet the current electricity demand. It is defined as:

_ critical ¼ m _ 4þm _ 18 m The thermodynamic analysis (i.e. first and second law analysis) of the system’s individual components will be presented in this section. In the analysis, the following assumptions will be adopted.  Steady state operation for all components.  The heat exchanger, pumps, compressor and turbines are adiabatic.

Fig. 2. Power management controller algorithm.

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Fig. 3. (a) High penetration configuration (energy analysis). (b) High penetration configuration (exergy analysis).

 Kinetic and potential exergetic terms are negligible for all system components except the WT.  The chemical exergetic term does not change in the turbine, pumps, compressor or the heat exchanger.  The ambient temperature and pressure are constant (T0 ¼ 298 K and P0 ¼ 100 kPa).  Molar flow rates for streams where any chemical reaction occur can be used in order to find the energy balance equations.  Air is treated as an ideal gas with a molar composition of 21% oxygen and 79% nitrogen.  The motor efficiency for the pump and compressor is assumed as 85%. A WT with a diameter of 100 m and betz limit (4) of 0.4 is chosen for the present analysis. The balance equations of each component for mass, energy, entropy and exergy are written as follows.

Fig. 5. (a) Wind addition configuration (energy analysis). (b) Wind addition configuration (exergy analysis).

 Wind Turbine (WT) (1e2)

_1 ¼ m _2 m _ 1 h1 þ m

(1a) V12 2

! _ 2 h2 þ ¼ m

V22 2

! _ WT þW

_ 2 s2 _ 1 s1 þ S_ g;WT ¼ m m _ 1 ex1 þ m

V12 2

! _ 2 ex2 þ ¼ m

(1b)

(1c) V22 2

! _ _ WT þ Exd þW WT

(1d)

 Compressor (3e4e5)

_3 ¼ m _ 4þm _5 m

(2a)

_ _ 4 h4 þ m _ 5 h5 _ 3 h3 þ W m Comp ¼ m

(2b)

_ 4 s4 þ m _ 5 s5 _ 3 s3 þ S_ g;Comp ¼ m m

(2c)

_ _ _ 4 ex4 þ m _ 5 ex5 þ Exd _ 3 ex3 þ W m Comp ¼ m Comp

(2d)

 Combustion Chamber (CC) (4e7e16e18)

Fig. 4. (a) Low penetration configuration (energy analysis). (b) Low penetration configuration (exergy analysis).

_ 4þm _ 16 þ m _ 18 ¼ m _7 m

(3a)

_ 4 h4 þ m _ 16 h16 þ m _ 18 h18 ¼ m _ 7 h7 m

(3b)

_ 4 s4 þ S_ g;CC þ m _ 16 s16 þ m _ 18 s18 ¼ m _ 7 s7 m

(3c)

_ _ 16 ex16 þ m _ 18 ex18 ¼ m _ 7 ex7 þ Exd _ 4 ex4 þ m m CC

(3d)

where

ex16 ¼ extf þ exch

(3e)

M. Rabbani et al. / Energy 44 (2012) 321e328

The chemical exergy of the flue gases at point 7 can be determined as:

exch

¼ RT0

X

yln

y ye

(3f)

325

_ 14 s14 þ S_ g;LPT ¼ m _ 15 s15 m

(8c)

_ _ _ 15 ex15 þ Exd _ 14 ex14 ¼ m m LPT þ W 5

(8d)

H OðgÞ

O2 2 2 where ye is given as yN ¼ 0:03 e ¼ 0:7567, ye ¼ 0:2035, ye CO2 and ye ¼ 0:0003. The total exergy of fuel can be expressed as:

ex16 ¼ hðT; PÞ  h0  T0 ðsðT; PÞ  s0 Þ þ RT0

X

yln

y ye

(3g)

 Gas turbine (GT) (7e8)

_7 ¼ m _8 m

(4a)

_ _ 7 h7 ¼ m _ 8 h8 þ W m 3

(4b)

_ 8 s8 _ 7 s7 þ S_ g;GT ¼ m m

(4c)

_ _ þm _ 7 ex7 ¼ W _ 8 ex8 þ Exd m 3 GT

 Condenser (15e14)

_ 15 ¼ m _ 10 m

(9a)

_ 10 h10 þ Q_ 2 _ 15 h15 ¼ m m

(9b)

Q_ _ 15 s15 þ S_ g;CD ¼ m _ 10 s10 þ 2 m T0

(9c) _

_ _ Q2 _ 10 ex10 þ Exd _ 15 ex15 ¼ m m CD þ Ex

(9d)

 Air cooler

(4d)

 Heat recovery heat exchanger (8e11e13e14e12e9)

_8 ¼ m _9 m

(5a)

_ 12 _ 11 ¼ m m

(5b)

_ 14 _ 13 ¼ m m

(5c)

_5 ¼ m _6 m

(10a)

_ 5 h5 ¼ m _ 6 h6 þ Q_ 1 m

(10b)

Q_ _ 5 s5 þ S_ g;CO ¼ m _ 6 s6 þ 1 m T0

(10c) Q_ 1

_ _ _ 5 ex5 ¼ m _ 6 ex6 þ Exd m CD þ Ex

(10d)

_ 11 h11 þ m _ 13 h13 ¼ m _ 9 h9 þ m _ 12 h12 þ m _ 14 h14 _ 8 h8 þ m m (5d) _ 8 s8 þ m _ 11 s11 þ m _ 13 s13 þ S_ g;HX ¼ m _ 9 s9 þ m _ 12 s12 þ m _ 14 s14 m (5e)

The net power input to the combined system when the compressor and pump are powered by the WT can be determined as:

(5f)

_ net;Input ¼ W _ þW _ ¼ h 4 W 1 2 m

_ 8 ex8 þm _ 11 ex11 þm _ 13 ex13 ¼ m _ 9 ex9 þm _ 12 ex12 þm _ 14 ex14 m _ þExd HX  Pump (10e11)

n X

_ WT W

(11a)

i¼1

_ 10 ¼ m _ 11 m

(6a)

_ P ¼ m _ 10 h10 þ W _ 11 h11 m

(6b)

_ 10 s10 þ S_ g;P ¼ m _ 11 s11 m

(6c)

_ net;Input ¼ W _ þW _ þ 1rAV 3 W 1 2 2

_ _ P ¼ m _ 10 ex10 þ W _ 11 ex11 þ Exd m P

(6d)

The net work rate output of the combined system when the compressor and pump are powered by the WT can be determined as:

_ 12 ¼ m _ 13 m

(7a)

_ _ 12 h12 ¼ m _ 13 h13 þ W m 4

_ þW _ þW _ 5 W 3 4 _ W net;Output ¼

(7b)

_ 12 s12 þ S_ g;HPT ¼ m _ 13 s13 m

(7c)

In the case the wind power addition configuration, the new work output

_ _ _ 12 ex12 ¼ m _ 13 ex13 þ Exd m HPT þ W 4

(7d)

_ _ _ _ _ _ _ W net;Output ¼ W 3 þ W 4 þ W 5 þ W WT  W 1  W 2

 High pressure turbine (HPT) (12e13)

In the present study, n ¼ 15 is chosen. In the case of wind power addition, the configuration input can be determined as

hG

The power input to the compressor is:  Low pressure turbine (LPT) (14e13)

_ 14 ¼ m _ 15 m

(8a)

_ _ 14 h14 ¼ m _ 15 h14 þ W m 5

(8b)

_ W ele;Comp _ W Comp ¼

hm

The power input to the pump is:

(11b)

(12a)

(12b)

M. Rabbani et al. / Energy 44 (2012) 321e328

Table 1 Thermodynamic state points. Temperature (K)

Pressure (kPa)

Mass flow rate (kg/s)

Enthalpy (kJ/kg)

Entropy (kJ/kg K)

Exergy (kJ/kg)

298 298 298 540.8 540.8 305 1773 926.8 523 318.9 325.2 773 570.6 773 318.9

100 100 100 1350 1350 1350 1350 300 300 10 6000 6000 1000 1000 10

4107 4107 697 633.6 63.36 63.36 633.6 633.6 633.6 76.11 76.11 76.11 76.11 76.11 76.11

298.4 298.4 298.4 545.5 545.5 305.5 1970 963.4 527.1 191.7 222.9 3422 3045 3478 1013

5.7 5.7 5.7 5.56 5.56 4.98 6.91 6.57 5.52 0.65 0.72 6.88 7.11 7.76 0.65

0 0 0 289.4 289.4 222.7 1312 405.5 281.3 2.85 10.78 1376 930.1 1170 824.5

160 140

Vc (m/s)

12

mc (kg/s)

120 100

10

Vc (m/s)

1 2 3 4 5 6 7 8 9 10 11 12 13 14 15

14

80 8

60

mc (kg/s)

326

40 6 20 4 40000

80000

120000

0 200000

160000

Load (kW) Fig. 6. Effect of load on critical parameters (based on plant parameters in Table 1).

_ _ P ¼ W ele;P W

0.8

hm

(13)

0.6

η

_ W net;Output ¼ _ net;Input _ 16 HHV þ W m

The overall exergy efficiency without the compressed air unit becomes:

hex

_ W net;Output ¼ _ net;Input _ 16 exch þ W m

(14)

__ _ W net;Output þ Q 2 _ _ HHV þ W m

0.7

0.4

0.6

0.3

0.5

0.2

0.4

10

hex;Cog

16

14

16

18

20

22

24

Fig. 7. Energy and exergy efficiency for high penetration configuration.

(15)

The overall exergy efficiency with cogeneration and without the compressed air unit is:

  To _ _ W net;Output þ Q 2 1  Ts ¼ _ net;Input _ ex þ W m

12

V (m/s)

net;Input

16

0.8

ηexe,C

0.5

The overall energy efficiency with cogeneration and without compressed air unit is expressed as:

hen;Cog ¼

0.9

ηexe ηen,C

ηc

0.7

The overall energy efficiency without compressed air unit becomes:

hen

1 ηen

(16)

ch

hex;H

  To _ _ W net;Output þ Q 2 1  Ts ¼ _ net;Input þ m _ 16 exch þ W _ 6 ex6 m

(20)

Table 1 shows the temperature, pressure, mass flow rates, enthalpy, entropy and exergy of the state points in the system.

The overall energy efficiency with cogeneration and compressed air storage for high penetration is expressed as:

__ _ _ 6 h6 W net;Output þ Q 2 þ m _ net;Input _ HHV þ W m

0.76

(17)

16

0.72

The overall exergy efficiency with cogeneration and compressed air storage for high penetration is expressed as:

(18)

hen;L ¼

_ W net;Output 2 _ _ _ 6 h6 þ m16 HHV þ W m net;Input

70

m4

60 50

0.68

40

0.66 30 0.64

The overall energy efficiency with cogeneration and compressed air storage for low penetration is expressed as:

_ þ Q_

ηen ηexe m18

0.7

η



hex;H

 To _ _ W 1  þ m_ 6 ex6 þ Q net;Output 2 Ts ¼ _ net;Input _ 16 exch þ W m

80

0.74

m (kg/s)

hen;H ¼

20 0.62 10

0.6

(19)

The overall exergy efficiency with cogeneration and compressed air storage for low penetration is expressed as:

0 2

4

6

8

10

V (m/s) Fig. 8. Energy, exergy efficiency and required mass flow rate from storage unit for low penetration configuration.

M. Rabbani et al. / Energy 44 (2012) 321e328

14

300

0.8

250

0.7

200

100

V,T=1300K V,T=1400K V,T=1500K V,T=1700K m,T=1300K m,T=1400K m,T=1500K m,T=1700K

8

6

η en

mc (kg/s)

0.5 0.5 0.4 0.4

50

0.3

0.3 0

0.2

4 0

50

100

150

200

0.2

0

250

5

10

Load (MW)

16

180 160

14

140 12

8

80

V

100

m (kg/s)

120 10

60 6 40 4

20

2

0 50

100

20

25

V (m/s)

Fig. 9. Effect of electricity demand on critical parameters at different combustion temperatures.

0

15

150

200

250

Load (MW) Fig. 10. Effect of electricity demand on critical parameters at different pressure ratios.

4. Results and discussion Based on the plant parameters (Table 1), Fig. 6 shows the effect of the load demand on the critical wind speed and the critical mass flow rate. Both parameters increase with an increase in the load (Eq. (11a)). Also the critical velocity and flow rate decrease with increases in the number of WT. For a 100 MW load demand and

Fig. 12. Energy and exergy efficiencies for wind power addition configuration.

plant parameters mentioned in Table 1, the critical velocity is calculated as Vc ¼ 10.4 m/s and the critical mass flow rate is mc ¼ 73.96 kg/s. If the wind speed increases from 10.4 m/s then the system is under high penetration and the mass flow of the air stored in the storage unit can be found by Eq. (2a). If the wind speed decreases from 10.4 m/s, then the system is under low penetration. Fig. 7 shows the effect of the wind speed under the high penetration configuration for a load demand of 100 MW on the energy and exergy efficiency with (Eqs. (17) and (18)) and without storage (Eqs. (15) and (16)). From the figure, the storage increases the efficiency (i.e. due to the extra term in the numerator of Eqs. (15) and (16)). With an increase in the wind speed, both the energy _ and exergy efficiency decreases due to the increase in W in (Eqs. (11a), (15) and (16)), however the mass flow of the compressed air also increases with an increase in the wind speed. Fig. 8 shows the effect of the wind speed under the low penetration configuration for a load demand of 100 MW on the energy and exergy efficiency (Eqs. (19) and (20)). At low wind speeds, the efficiency is higher due to less power input into the system. _ 6 is high but it has less However, at low wind speed, the value of m exergy potential (refer to enthalpy and exergy of point 6 in Table 1). With an increase in the wind speed, both the efficiencies and the compressed air flow from the storage tank decrease and the mass flow rate of the compressed air supplied by the compressor increases. Fig. 9 shows the variation of the load demand on the critical wind velocity and required mass flow rate across the CC at different

160

14

500

350

140 300

12

m4 (kg/s) m5 (kg/s) WWT (MW)

120 250

8

6

80 60

300 200

m (kg/s)

V, T=600K V, T=700K V, T=800K V, T=900K m, T=600K m, T=700K m, T=800K m, T=900K

m (kg/s)

100

10

V (m/s)

400

150

200

100 100

40 50

20

0 0

4

0 0

50

100

150

200

250

Load (MW)

0

5

10

15

20

25

30

Velocity (m/s) Fig. 11. Effect of electricity demand on critical parameters at different Rankine cycle temperatures.

Fig. 13. Energy and exergy efficiencies for wind power addition configuration.

Power Output From Wind Turbine (MW)

Vc (m/s)

150

0.6

0.6

12

10

0.7

η exe

16

327

328

M. Rabbani et al. / Energy 44 (2012) 321e328

combustion temperatures. Increasing the combustion temperature reduces the critical velocity and the critical mass flow rate across the CC for a given load demand. This is explained in terms of the energy balance across the GT (i.e. higher temperature difference across the GT yields more work output). Fig. 10 shows the variation of the load demand on the critical wind velocity and required mass flow rate across the CC at different pressure ratios. For a given load demand, increasing the pressure ratio increases the work input to the compressor (Eq. (2b)) hence the value of the critical velocity also increases which results in a decrease in efficiency (see Figs. 6 and 7). However it does not significantly affect the critical mass flow rate and a small variation is observed at the high load demand. Fig. 11 shows the variation of the load demand on the critical wind velocity and required mass flow rate across the CC at different boiler temperatures in the Rankine cycle. A variation in the boiler temperature does not significantly affect the critical velocity for a given load demand because the Rankine cycle is a secondary cycle and the required work by the pump is far less than the compressor. In this analysis, the mass flow in the Rankine cycle is found by applying an energy balance across the HRSG. Fig. 12 shows the effect of wind speed on the energy and exergy efficiency in the wind addition configuration (Fig. 4). The net output work in the wind addition configuration is calculated by using Eq. (12b). Increasing the wind speed reduces the energy and exergy efficiency because the net energy input to the system increases. However, if the wind speed is greater than the critical velocity, then the wind power addition is under high penetration. With an increase in the wind speed, the mass flow rate of the air in the _ 4 Þ for the wind power configuration. Brayton cycle decreases ðm Fig. 13 shows the effect of a variation of wind speed on the mass flow rate into the CC and storage unit and power output from the WT. When all of the load demand is supplied by the WT, then at _ 4 Þ and mass flow rate that point, the mass flow across the CC ðm _ 5 Þ is zero. Additional wind power is used input to the storage unit ðm for compressed air storage. 5. Conclusions In this paper, a WT was coupled with a combined cycle to increase the net power output of the plant. Three different WTs and combined cycle integrated configurations were analyzed. The load demand affects the critical velocity and critical mass flow rate. In the case of low demand and high penetration, extra wind power was used to compress air which can be used during a low penetration configuration. During high load demand, all of the wind power is used to drive the pump and compressor, and if required, additional compressed air is supplied by the storage unit. The

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