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International Journal of Emerging Technology and Advanced Engineering

Website: www.ijetae.com (ISSN 2250-2459, ISO 9001:2008 Certified Journal, Volume 4, Issue 2, February 2014)

612

Parametric Analysis of Combined Cycle Power Plant

Using Inlet Vapour Compression Cycle

L. J. Nayak

1

, D. Mahto

2 1

ME student, 2Assistant Professor, Department of Mechanical Engineering, BIT, Mesra, Jharkhand 835217, India

Abstract—The study deals with combined cycle power plant

(single pressure HRSG) with inlet air cooling. Combined cycle power plant are being extensively used in view of their capability of offering higher specific work output and efficiency compared to other thermal power plant for same fuel consumption. Combined cycle performance is strongly affected by climate condition. During hot days power output is reduced. Cooling the inlet air can increase the net power output. This is because decrease the inlet air temperature will increase its density causing higher mass flow rate. So vapour compression refrigeration system (mechanical refrigeration system) at inlet of compressor is used, which reduce the temperature of inlet air to desired level. Power output and cycle efficiency are simulating with respect to the cycle temperature and pressure ratio for a typical set of operating condition.

Keywordshot climate, inlet air cooling, vapour compression cycle, gas turbine, combined cycle power plant.

I. INTRODUCTION

Power output is decreases with increase in inlet air temperature to compressor. As specific volume of air is directly proportional to temperature, reduction in inlet air temperature results in increase in density of air and increase in air density results in a higher air mass flow rate, once the volumetric rate is constant.Consiquently the combined cycle gas turbine power plant enhances. There are different type of cooling methods are used in cooling inlet air such as evaporative cooling, high pressure fogging, mechanical

refrigeration (vapour compression refrigeration),

absorption cooling etc. Mechanical refrigeration is one of the efficient method which is operating in hot and humid climate. This system is not sensitive to ambient air wet bulb

temperature. Vapour compression refrigeration system

have no limitation on achievable inlet air temperature and duration of inlet air cooling.

M.M. Alhazmy et al.[1] investigated the performance of gas turbine power plants by cooling the inlet air to compressor. A comparison between two types of air coolers such as water spraying system and cooling coil is performed.

The performance characteristics are examined for a set of design and operational parameters including ambient

temperature, relative humidity, and turbine inlet

temperature and pressure ratio.

The results show that spray coolers are capable of boosting the power and enhancing the efficiency of the gas turbine power plant in a way that is much cheaper than the cooling coils, however, it operates more efficiently at hot and dry climates. Cooling coils give a full control on the compressor inlet conditions; however they consume considerable amount of power.

Rahim K. Jassim et al.[2] reported that Gas turbine (GT) power plants operating in arid climates suffer a decrease in output power during the hot summer months because of insufficient cooling. An energy analysis of a GT Brayton cycle coupled to a refrigeration cycle shows a promise for increasing the output power. Result shows that the enhancement in output power depends on the degree of chilling the air intake to the compressor (a 12 - 22 K decrease is achieved). performance analysis of the a GT, for a pressure ratio of 10, rate of air intake of 250m3/s and 1000oC maximum cycle temperature shows that the intake air temperature decreases by 12 and 22 K, while the PGR increases a maximum of 15.46%. The average increase in the plant power output power is 12.25%, with insignificant change in plant thermal efficiency.

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International Journal of Emerging Technology and Advanced Engineering

Website: www.ijetae.com (ISSN 2250-2459, ISO 9001:2008 Certified Journal, Volume 4, Issue 2, February 2014)

613

II. MODELING AND GOVERNING EQUATION

A. Vapour Compression Cycle Modeling:

A simple analysis of standard vapour compression refrigeration system can be carried out By assuming:

-

Steady flow process through each component.

-

Negligible kinetic and potential energy changes.

-

No heat transfer and pressure drops in connecting pipe lines.

Fig.1. T S diagram of vapour compression cycle

COP

=

Work input to vapour compression cycle,

Wvcc =

( )

(1)

Where Ta= ambient temperature of air

T1=inlet temperature of air to compressor

B. Simple Combined Cycle:

B.1. Gas modeling:

Specific heat of air at constant pressure has been assumed to be function of temperature only Specific heat of air is taken for two range of temperature, given as below:

For air at low temperature range of 200-800K,

=1.0189× -0.13784 +1.9843× +

4.2399× – 3.7632 × (2)

For air at higher temperature range of 800-2200K,

=7.9865× +0.5339 –2.2882× +

3.7421× (3)

B.2. Compressor modeling:

The compressor Used in gas turbine power plant is of axial flow type. The thermodynamic losses. In an axial flow compressor are incorporated in the model by introducing the concept of polytropic efficiency.

For a specified inlet and polytropic efficiency, the temperature is obtained by integrating,

=

[

]

(4)

There is approximately equal pressure ratio in each stage and the number of stage (zc) is calculated by the relation

rp, stage = (rp,c)

1/zc

(5)

Fig.2.schematic diagram of compressor

Mass balance equation for compressor,

̇i = ̇e (6)

Energy balance Equation for compressor

Wc = ̇e - ̇i (7)

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International Journal of Emerging Technology and Advanced Engineering

Website: www.ijetae.com (ISSN 2250-2459, ISO 9001:2008 Certified Journal, Volume 4, Issue 2, February 2014)

614

Fig.3. Schematic diagram of simple combined cycle with inlet vapour compression cycle

B.3.Combution chamber modeling:

Natural gas is burnt in the presence of compressor air from compressor and here after the working fluid of the cycle changes from air to gas. The specific heat of gas at constant pressure has been adopted from [4].

Fig.4. schematic model of combustion chamber

Mass balance equation for combustion chamber,

̇e = ̇i+ ̇f (8)

Energy Balance equation for combustion chamber,

̇i + ̇f (LHV)f = ̇e (9)

B.4.Gas Turbine modeling:

To evaluate the gas turbine work output, the following assumptions and expressions are used. The expansion process in each row is assumed to be polytropic and is divided into six part. Each row of the turbine treated as expander whose walls continuously extract work. The temperature at a state in turbine is determined by the relation

=

[{

}

]

(10)

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International Journal of Emerging Technology and Advanced Engineering

Website: www.ijetae.com (ISSN 2250-2459, ISO 9001:2008 Certified Journal, Volume 4, Issue 2, February 2014)

615

Mass balance equation for gas turbine,

̇i = ̇e (11)

The gas turbine work is the sum of the work done by all rows of bladings,

∑ ̇g,i( ) (12)

Where ‘i’ and’e’denotes the inlet and exit condition, receptively of a row.

B.5. HRSG modeling:

An unfired HRSG (single pressure)has been considered here. Temperature and pressure of steam depends upon the temperature of gas turbine exhaust entering into HRSG.

The HRSG consists of three heat exchanger section: economizer, evaporator, superheater.

In the economizer, the feed water is heated up to a temperature to its saturation point. In the evaporator, feed water is evaporated at constant temperature and pressure. The water and saturated steam are separated in drum and steam is fed to super heater where it is superheated to the desired live steam temperature.

Fig.6. T-Q diagram of simple gas turbine cycle using Single pressure HRSG

Pinch point temperature is the difference between the evaporator temperature on the water or steam side and outlet temperature on the exhaust gas side.

Tpp = Tg,b - Ts

Where Tpp = pinch point temperature difference

Tg,b = Temperature of gas at point b

Heat available with exhaust gases from gas turbine can be given as,

Qav = ̇g ( )× (13)

Where = Exhaust temperature of gas turbine = Stack temperature

= heat loss factor

Heat supplied by superheater ,

Qsh = ̇s( ) (14)

Where enthalpy of superheated steam

= enthalpy of saturated steam

Energy balance equation for HRSG,

̇s( )= ̇g ( )× (15)

B.6. Steam turbine modeling:

Fig.7. schematic of steam tubine

̇s = mass rate of steam generated in HRSG

̇s1 = mass rate of steam bled for deaerator

Mass and energy balance yield steam turbine output given by

= ∑ ̇si( ) (16)

Where ( )= ( )isentropi

̇si = amount of steam entering to the respective main

turbine stages as per configuration.

B.7. Deaerator modeling:

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International Journal of Emerging Technology and Advanced Engineering

Website: www.ijetae.com (ISSN 2250-2459, ISO 9001:2008 Certified Journal, Volume 4, Issue 2, February 2014)

616

Mass balance equation is,

̇s,d,i +( ̇s - ̇s,d,i ) = ̇s,d,e (17)

Energy balance equation is,

̇s,d,i + ( ̇s - ̇s,d,i ) = ̇s (18)

B.8. Feed pump modeling:

The mass and energy balance give the pump work input as below

= ∑ ( ) (19)

Where i and e denotes the inlet and exit condition of pump.

B.9.Performance parameters:

Net work done by gas turbine cycle,

Wn,gtc = (Wgtηm)– η (20)

Net work done by steam cycle,

Wn,stc = (Wstηm) – η (21)

Efficiency of combined cycle,

ηcc =

(22)

Where Qa is heat supplied to the combustion chamber.

Qa = ̇f (LHV)ηcomb

Plant Specific work,

Wsp=

(23)

III. INPUT DATA FOR ANALYSIS [5]

Parameter Symbol unit

Compressor (i) polytropic efficiency (ηp,c)=92.0 %

(ii)mechanical efficiency(ηm,c)=98.5 %

Combuster (i) combuster efficiency (ηcomb) =99.5 %

(ii) Pressure loss (ploss) =2.0% of entry %

Pressure

(iii) Lower heating value (LHV)=42 MJ/kg

Gas turbine (i)polytropic efficiency(ηp,t)=92.0 %

(ii)exhaust pressure=1.05 bar

HRSG (single (i) pressure=50 bar Pressure) (ii) pinch point temperature K Difference =10

(iii) Stack (minimum temperature)=363 K (iv)Max. Superheat temperature=843 K (v) Deaerator pressure=1.2 bar (vi) Condenser pressure=0.05 bar

Steam turbine (i) Isentropic efficiency=88 % (ii)Mechanical efficiency=98.5 % (iii)Minimum steam quality at exhaust=88 dry Alternator Alternator efficiency=98.5 %

Nomenclature:

Cp Specific heat (kJ kg-1 k-1)

W Specific work (kJ kg-1) R Gas constant (kJ kg-1 k-1) h Specific enthalpy (kJ kg-1) T Temperature (k)

̇ mass flow rate(kg s-1) η efficiency(%) υ specific volume(m3

kg-1) p pressure(bar)

Subscripts:

a air, ambient c compressor cc combined cycle comb combustion chamber d deaerator

i inlet e exit f fuel g gas gt gas turbine gtc gas turbine cycle m mechanical n net

s steam st steam turbine

stc steam turbine cycle vcc vapour compression cycle w water

sh superheated steam p polytropic

Acronyms:

COP coefficient of performance LHV lower heating value

HRSG heat recovery steam generator GTIT gas turbine inlet temperature CC combustion chamber

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International Journal of Emerging Technology and Advanced Engineering

Website: www.ijetae.com (ISSN 2250-2459, ISO 9001:2008 Certified Journal, Volume 4, Issue 2, February 2014)

617

IV. RESULT AND DISCUSSION

285 290 295 300 305 310 315

450 460 470 480 490

C

o

mp

re

sso

r

w

o

rk(kJ/

kg

)

compressor inlet Temperature(K) rpc=20

GTIT=1700K

Fig.9. Effect of compressor inlet temperature on compressor work

Fig.9 show that as inlet temperature of air to compressor increases, work input to compressor increase. So to decrease the compressor load, inlet air cooling is necessary.

12 14 16 18 20 22 24

500 550 600 650 700 750 800

Pl

a

n

t

sp

e

ci

fi

c

w

o

rk(kJ/

kg

)

Compressor pressure ratio GTIT=1700K

Fig.10.Effect of rpc on Plant specific work

Fig.10 shows the effect of variation of compressor pressure ratio on specific work. Specific work slightly decreases with increase in pressure ratio. So, for higher value of specific work, lower pressure ratio is required.

12 14 16 18 20 22 24

50 51 52 53 54 55

Pl

a

n

t

e

ff

ici

e

n

cy(%

)

Compressor pressure ratio GTIT=1700K

Fig.11.Effect of rpc on Plant efficiency

Fig.11 shows the effect of combined cycle efficiency with compressor ratio for given GTIT. Plant efficiency increases with increase in compressor pressure ratio, but the rate of increase is very low.

1100 1200 1300 1400 1500 1600 1700

100 200 300 400 500 600 700

Pl

a

n

t

sp

e

ci

fic

w

o

rk(kJ/

kg

)

Turbine inlet temperature(k) Compressor pressure ratio=20

Fig.12.Effect of GTIT on Plant specific work

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International Journal of Emerging Technology and Advanced Engineering

Website: www.ijetae.com (ISSN 2250-2459, ISO 9001:2008 Certified Journal, Volume 4, Issue 2, February 2014)

618

1100 1200 1300 1400 1500 1600 1700

30 35 40 45 50 55

Pl

a

n

t

e

ff

ici

e

n

cy(%

)

Turbine inlet temperature(k) compressor pressure ratio=20

Fig.13.Effect of GTIT on Plant efficiency

Fig.13 shows the effect of combined cycle efficiency with GTIT at a given value of pressure ratio. The efficiency tends to increases with increase in GTIT.

The fig.14 illustrates the advantage of inlet air cooling combined cycle over the simple combined cycle without inlet air cooling.It is observed that specific work of inlet air-cooled combined cycle is higher than that without inlet air cooling for a given ambient temperature. For a given relative humidity, as the ambient temperature increases mass flow rate of air decreases and compressor load is increased.

0 100 200 300 400 500 600 700

Pl

a

n

t

sp

e

ci

fi

c

w

o

rk(kJ/

kg

)

without cooling inlet air cooling inlet air to 150 c rpc=20,GTIT=1700K

atmospheric temperature=400 c

Fig.14

V. CONCLUSION

The present model involved in thermodynamic analysis of gas and steam turbine based combined cycle using inlet air cooling. Performance curves shows that higher the GTIT, higher is the efficiency and specific work. Inlet air cooling by vapour compression cycle can increase the specific work output of combined cycle power plant.

REFERENCES

[1] Alhazmy, M.M.,Najjar,Y.S.H. ,Augmentation Of gas turbine

performance using air coolers, Applied thermal Engineering

24,2004.

[2] Jassim,R.K., Zaki,G.M.,Alhazmy,M.M.,Thermo- Economic analysis

of a Gas turbine power plant With cooled air intake, Yanbu Journal of Engineering and Science Vo11, Oct.2010.

[3] Ondryas, I.S.,Wilson,D.A., Kawamoto, M., Haub G.L., Option in

gas Turbine power augmentation using inlet air chilling,engineering gas Turbine and power, ASME113 (1991) 203-2012.

[4] Sanjay, energy and exergy analysis of combined cycle systems with

different bottoming cycle configurations, Int.J.Energy Res.37:899-912,2013.

[5] Sanjay,Onkar Singh,B N Prasad,Infuence of different means of

Turbine blade cooling on thermodynamic performance of combined cycle,Applied thermal Engineering 28(2008) 2315-2326.

[6] Nag P.K. (2009). Power Plant Engineering. Tata McGraw-Hill

Publishingcompany limited.New Delhi.ISBN-13:978-0-07-064815-9.

[7] Kehlhofer R. (1997). Combined cycle Gas and steam Turbine

power Plant .pennwell publishing company.ISBN-0-87814-736-5.

[8] Yadav R (2004) steam and gas turbine and Power plant Engineering,

Central publishing House,Allahabad.

[9] Santos, A.P., Analysis of gas Turbine Performance with inlet air

Cooling techniques , appliedtoBraziliansites,J.Aerospace. Technol Manag. Vol.4. No.3,pp.314-353,jul-sep.2012.

[10] Kakaras,E., Doukelis,A.,Prelipceanu, A.,Karellas,S., Inlet air cooling

Methods for gas turbine based power plants,Journal of Engineering for GasTurbines and Power,(2006),vol128,312-317.

References

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