EXERGY ANALYSIS OF A COMBINED GAS/ STEAM TURBINE

CYCLE WITH A SUPERCHARGED BOILER

S. A. Abdel-Moneim and . K. M. Hossin *

* Graduate student, Mech. Eng. Dept. Faculty of Eng. Garyounis Univ., Libya.

Authors’ contributions This work was carried out in collaboration between both authors. Author SAAM designed the study and author KMH was M Sc research student who carried out simulation code and computations. Author SAAM analyzed the data, discussed the results and wrote the final manuscript, Both authors read and approved the final manuscript.

ABSTRACT

In this paper, energy and exergy analysis of a supercharged boiler combined cycle (SBCC) was carried out. A combination of a supercharged boiler (SB) and a heat recovery boiler (HRB) was investigated. The effects of the inlet temperature of the gas turbine, the excess air factor, and the compressor pressure ratio on the performance of the SBCC were studied. Comparisons between the SBCC and the conventional combined cycle were performed. The results indicated that the SBCC gives output power up to 2.1 times of that of the conventional combined cycle when compared at the same values of the operating parameters. However, the SBCC efficiency was found to be lower than the conventional combined cycle. The exergy analysis showed an advantage of SBCC over the conventional combined cycle. Keywords: Thermal power plant; supercharged boiler, combined cycle, energy; exergy;

second-law efficienc, exergy destruction. NOMENCLATURE

PC specific heat at constant pressure, (kJ/kmol K) e flow specific exergy, (kJ/kg) ExD exergy destruction rate, (kW) h enthalpy, (kJ/kg)

h enthalpy difference, (kJ/kg) LHV lower heating value of fuel, (kJ/kmol) m mass flow rate, (kg/s)

HPm mass of HP steam generated in HRB, (kg/kmoln.g)

LPm mass of LP steam generated in HRB, (kg/kmoln.g)

LHm mass ratio of LP to HP steam

SBm mass of steam generated in SB, (kg/kmoln.g) M molecular weight; (kg/kmol) P pressure, (bar) or power, (kW)

PR ST to GT power ratio

q heat transferred per kg of steam, (kJ/kg) T temperature, (K) v specific volume, (m3/kg) w work per kg of steam, (kJ/kg) W work, (kJ/kmoln.g) Xa actual air to fuel ratio, (kmola/kmoln.g) Xg amount of product gases, (kmolg/kmolen.g)

Greek symbols

321 ,, steam mass fractions

HRB Efficiency of the HRB, (%)

SB Efficiency of the SB, (%)

com thermal efficiency of combined cycle, (%)

G generator efficiency, (%)

CG thermal efficiency of GT cycle, (%)

HRB thermal efficiency of HRB steam cycle, (%)

m mechanical efficiency, (%)

P pump isentropic efficiency, (%)

SB thermal efficiency of SB steam cycle, (%)

nd2 second-law efficiency, (%) excess air factor

C compressor presser ratio Subscripts a air com combined cycle g product gases GC gas turbine cycle i inlet

n.g natural gas N net

o outlet SC steam turbine cycle Acronyms C compressor CON condenser CP condensate pump, circulated pump EC economizer EV evaporator FP feed pump FWH surface feed-water heater

GEN generator GT gas turbine HRB heat recovery boiler HP high pressure LP low pressure NG natural gas P pump SB supercharged boiler SBCC supercharged boiler combined cycle SH superheater ST steam turbine

1. Introduction

Nowadays combined gas/steam turbine cycle power plants are widely used for cogeneration of electricity and process steam. Examples are combined cycle power plants coupled with sea-water desalination, district heating plants, chemical industries, etc. In combined heat and power plants, the gas turbine exhaust heat is utilized through the use of heat recovery steam generators or boilers (HRBs). The main methods of improving the overall efficiency of combined power plants are: increasing the mean temperature of heat supplied by increasing the inlet gas temperature of the gas turbine and decreasing the mean temperature at which heat is rejected. Moreover, optimization of the plant components and its operating parameters is an important method for improving its efficiency. For example, any changes made in the HRB operating parameters (i.e. the pinch point, approach temperature, first and second stage pressures, and mass ratios) can greatly affect the HRB performance and will eventually affect the overall combined plant performance. Briesch et al. [1] reported that 60 % efficiency is achievable by increasing the gas turbine inlet temperature to 1427ºC and by incorporating of advanced cooling techniques in the gas turbine expander. Utilizing reheat gas and steam turbines and increasing the reheat pressure and temperature, by applying steam cooling to the gas turbine blades, increase the average temperature of heat addition, thus, increasing the overall cycle efficiency [2,3,4]. Modeling and optimizing a dual pressure reheat combined cycle was carried out by Bassily [5] with introducing a technique to reduce the irreversibility of the steam generator. Recently, in an attempt to conserve energy, special attention is being paid to combined cycle power plants. One of the applicable methods of saving energy, in addition to reducing steam generator size, is to supercharge the steam generator by using a gas turbine-driven compressor to furnish combustion air. This new steam generator is called supercharged boiler (SB). Developments in materials science (metallurgy) and pressure vessel technology make it possible to build such a steam generator. The reduction in size and heat transfer surface of a supercharged boiler is due to two reasons. First, as the operating gas-side pressure is increased, the emissivity of the nonluminous radiating gases increases markedly. Second, the higher gas density and available pressure drop permit much higher gas mass flow rates (compared with the conventional steam generator) to be used in the convection section, with higher accompanying convection heat transfer coefficients [6]. Mikhael et al. [7] investigated the possibility of utilizing the solar energy for electrical power generation with a hybrid mode of steam generation in a combined power plant incorporating a SB and a HRB. Traditional first-law cycle analysis based upon component performance characteristics coupled with energy balances invariably lead to a correct final answer. However, such

analysis cannot locate and quantify the sources of loss that lead to that result. This is because the first law embodies no distinction between work and heat, no provision for quantifying the quality of energy [8]. On the other hand, the exergy analysis identifies the

location, the magnitude and the sources of thermodynamic inefficiencies in a thermal system. This information, which cannot be provided by other means (e.g., an energy analysis), is very useful for improving the overall efficiency and cost-effectiveness of a system or for comparing the performance of various systems [9]. The literature survey reveals that most of the studied combined gas/steam power plants use conventional steam generators (HRBs). The combined power plants incorporating supercharged boilers were not extensively investigated. In this paper, a supercharged boiler combined cycle (SBCC) is modeled and analyzed. 2. Description of the Cycle

Figure 1 shows the supercharged boiler combined cycle (SBCC) which combining the supercharged boiler cycle with the heat recovery cycle. In this combined cycle, the compressor supplies pressurized air to the SB (state 2). All combustion takes place in the boiler and steam can be generated at any suitable pressure and temperature (state 20S). The steam generated in the SB is circulated through a separated steam cycle. The steam expands in a steam turbine (ST) with extracted steam fractions during expansion process to heat the water before entering the SB. High-temperature pressurized gas from the boiler is expanded as it flows through the gas turbine (GT). The power so developed supplies the compressor and drives the generator. The hot exhaust gases from the GT pass through a dual pressure HRB to generate steam and next go to the stack (state 9). After the water leaves the condenser (state 1S), it is pumped to the dual pressure HRB, where it is converted to a steam with low and high pressures (states 9S and 7S, respectively). The low pressure steam is mixed with the exhaust steam from the high pressure turbine (HPST) before entering the low pressure turbine (LPST) to expand to the condenser pressure (state 11S). The T-Q and T-s diagrams of the HRB steam cycle are illustrated in figures 2 and 3, respectively. The T-s diagram for the SB steam cycle is shown in Fig. 4.

7s

6s 5s

9s

4s

3s

2s

T

Q

4 5

6 7

8 9

Fig. 2 T-Q diagram for the dual-pressure HRB.

1

2 3

4

5

6

7

8

9

1s

2s 3s

4s 5s

6s

7s 8s

9s

10s

11s

LP EC

LP EV

HP EC

HP EV

HP SH

C GT

HP ST LP ST

CON 1

LPP

HPP

GEN

CP

CP DRUM

Fig. 1 Schematic diagram for the SBCC.

SB ST

24s

12s 13s 15s 14s 16s 17s 18s

19s 21s

22s 23s

20s

CP

FP

FWH1

CON 2

GEN

GEN

To Stack

Air

P1 P2

HRB

μ1 μ2 μ3

1- μ1-μ2-μ3

FWH2 FWH3

Fuel NG

3. Thermal Analysis

In order to evaluate the performance of a combined cycle, thermal analysis of each component must be carried out. Appropriate assumptions are made through analyzing the combined cycle. These assumptions are:

- The changes of environmental or technological factors, such as ambient conditions and the efficiencies of the cycle components are not considered.

- Temperature differences and pressure drop through gas and steam pipes are assumed to be negligible.

- The influence of heat losses and pressure drop for feed-water heaters and condensers are not considered.

- The steam side pressure drop in HRB and SB are negligible. - Air leakage through gas cycle components can be neglected. The input data and other assumptions used in the present study are listed in Table 1.

Fig. 4 T-s diagram for the SB steam cycle.

s

T

PSB

PH3

PCON

PH2

PH1

12s

13s 14s

15s 16s

17s 18s

19s

20s

21s

22s

23s

24s

Fig. 3 T-s diagram for the dual-pressure HRB steam cycle.

s

8s PLP

1s

5s

9s

11s

6s

2s 3s

4s

10s

7s T

PHP

PCON

Before analyzing the performance of the combined cycle adopted, various operating parameters are varied to study their influences. In this study, three gas turbine inlet temperatures (T3) are studied (1200ºC, 1300ºC, and 1400ºC) and the excess air factor (λ) for the SB is changed from 1.2 to 2.2. Furthermore, this study covers a range of compressor pressure ratios (πc) from 6 to 30.

3.1 Analysis of the GT cycle

Referring to the schematic T-s diagram in Fig. 5, it is assumed that the GT cycle operates according to actual Brayton cycle. It consists of three main processes as follows: 3.1.1 The compression process

In this process, ambient air is drawn into the compressor at state 1, where its temperature and pressure are raised to state 2. The work absorbed by the compressor per kmol of air is determined from the following equation: )( 12, TTCw aPC kJ/kmolair (1) where aPC , is calculated at the average temperature between inlet and outlet of the compressor.

3.1.2 The combustion process

Through the combustion process, the fuel is burned to release the energy which is transformed into electricity by the GT, and by using the HRB the waste heat energy at GT outlet can generate steam which drives a ST to produce electric power. The fuel considered in the present study is the clean natural gas of ultimate analysis of 78.8 % methane (CH4), 14 % ethane (C2H6), 6.8 % nitrogen (N2) and 0.4 % carbon dioxide (CO2) by volume [6]. The combustion calculations will be based on 1 kmol natural gas.

T

s

2

1

3

4 4

2

9

Pa

P2

Fig. 5 T-s diagram for GT cycle.

Ta

Table 1 Assumptions of the cycle.

Assumption Unit Const. used Air mass flow rate Ambient temperature. Atmospheric pressure. Compressor isentropic efficiency.*

GT isentropic efficiency.*

Gas-side pressure loss in SB.*

Efficiency of SB. GT exhaust gas pressure. Pump isentropic efficiency.*

ST isentropic efficiency.*

Condenser pressure. Efficiency of HRB. Pinch point of HRB at HP.*

Pinch point of HRB at LP.*

LP to HP steam mass ratio. Mechanical efficiency.*

Generator efficiency*

kg/s oC bar % % % % bar % % bar % oC oC - % %

67.9268 30

1.01325 85 90 6 95

1.05 70 87

0.075 95 15 25 0.2 99 98

* Akiba and Thani [6]

To calculate the quantity of air required to burn 1 kmol of fuel gas, the combustion analysis is considered. The analysis is conducted as follows: - The combustion equation of the natural gas is written as:

2222

2222624

068.0.76.3).1(996.1072.1

)76.3(.068.0004.014.0788.0

22

2

NnOnOHCO

NOnNCOHCCH

OO

O

(2)

where λ is the excess air factor and 2On is the theoretical O2 required to burn 1 kmol of

natural gas ( 066.2004.02/996.1072.12

On kmol/kmoln.g.) - Energy balance equation for the combustion process based on 1 kmol of fuel is written as: SBioSBgPggngnPaPa hhmTCXTCLHVTCX 3,....,2, (3) LHV is the lower calorific value of the natural gas which given by:

626244. HCHCCHCH LHVnLHVnLHV kJ/kmoln.g. (4)

aX is the actual amount of air (number of kmoles) per kmol of fuel, and gX is the amount

of product gases per kmol fuel. The mass flow rates of the fuel and combustion gases are then calculated from the mass

flow rate of the air as follows: )//( .. gnaaagn MMXmm kg/s

gngnggg mMMXm ../ kg/s (5)

3.1.3 The expansion process

In this process, the outlet combustion gases from combustion process enter the gas turbine at a temperature of (T3). The gases then expand adiabatically to a temperature of (T4) which is taken as the temperature of the flue gases at the end of the expansion process. The work done by the GT per kmol natural gas is determined from the following equation:

)( 43, TTCXW gPgGT kJ/kmoln.g. (6)

here gPC , is also determined at the average temperature between inlet and outlet of the GT. The net work for the GT cycle is given by:

GmCamGTGCN wXWW /, kJ/kmoln.g. (7) The thermal efficiency of the GT cycle is calculated as follows:

gngnP

GCNGC TCLHV

W

..,

,

(8)

3.2 Analysis of the ST cycles

It is clear that this combined cycle consists of two separated steam cycles, Fig. 1. One of these is a HRB cycle and the other is a SB cycle. The ST cycles operate according to the Rankine cycle. The thermodynamic properties of water and steam are calculated point by point by using certain subroutines and from energy balances. The steam conditions used in this cycle are shown in Table 2. According to the numbers of different points in the schematic sketch and the corresponding T-s diagrams of the cycle, the following analysis is considered:

3.2.1 Analysis of the HRB steam cycle

Enthalpy rise in each pump in the cycle is written as: pioiP PPvh / kJ/kg (9)

where P is the pressure in (KPa) The heat added to the steam in each stage of the HRB is: - Low-Pressure Economizer ))(( 23 ss

hhmmQ LPHPLPEC kJ/s or ))(1( 23 ss

hhmq LHLPEC kJ/kgHP (10)

where mLH is the mass ratio of LP to HP steam in the HRB; HP

LPLH m

mm

- Low-Pressure Evaporator )( 39 sshhmq LHLPEV kJ/kgHP (11)

- High-Pressure Economizer ss

hhqHPEC 45 kJ/kgHP (12) - High-Pressure Evaporator

sshhqHPEV 56 kJ/kgHP (13)

- High-Pressure Superheater ss

hhqHPSH 67 kJ/kgHP (14) Total heat added in the HRB per kg of HP steam is then:

HPSHHPEVHPECLPEVLPECHRB qqqqqq kJ/kgHP (15) The work of HP and LP pumps per kg of HP steam are: HPPHPP hw kJ/kgHP (16) LPPLHLPP hmw 1 kJ/kgHP (17) The work of HP and LP stages of the ST per kg of HP steam are:

sshhwHPST 87 kJ/kgHP (18)

ss

hhmw LHLPST 11101 kJ/kgHP (19) The net work of HRB steam cycle per kg of HP steam is given by: GmLPPHPPmLPSTHPSTHRBN wwwww /, kJ/kgHP (20)

The thermal efficiency of HRB steam cycle is calculated as follows:

HRB

SCNHRB q

w , (21)

It is obvious that the above equations are based on kg of HP steam. To calculate the mass of HP steam, energy balance between points 4 and 8 in the HRB should be carried out:

LPEVHPECHPEVHPSH

gPgHRBHP qqqq

TTCXm

)( 84, kg/kmoln.g (22)

where T8 is determined using the temperature difference at LP pinch point ( LPPPT , ) as: LPPPLPsat TPTT ,8 )( (23) Then, the net work of HRB steam cycle per kmol natural gas is equal to: HRBNHPHRBN wmW ,, kJ/kmoln.g (24) 3.2.2 Analysis of the SB steam cycle

In order to calculate the fraction of steam required for each surface heater, energy balances for the surface heaters are done. - Energy balance for surface feed-water heater 1:

ss

ss

hhhh

1721

17181

(25)

- Energy balance for surface feed-water heater 2:

ss

ss

hhhh

1522

151612

))(1(

(26)

- Energy balance for surface feed-water heater 3:

ss

ss

hhhh

1323

1314213

))(1(

(27)

The work of the cycle pumps per kg of steam is: - Feed pump FPFP hw kJ/kg (28) - Heater pump 1 )1( 111 PP hw kJ/kg (29) - Heater pump 2 )1( 2122 PP hw kJ/kg (30) - Condensate pump )1( 321 CPCP hw kJ/kg (31) Specific work of ST for SB steam cycle (per kg of steam) is:

))(1()(

)1())(1(

24233212322

21222112120,

ssss

ssss

hhhh

hhhhw SBST

kJ/kg (32)

Net work of SB steam cycle per kg of steam is: GmCPPPFBmSBSTSBN wwwwww 21,, kJ/kg (33) Heat added in the SB per kg of steam is given by:

SShhqSB 1920 kJ/kg (34)

Thermal efficiency of the SB steam cycle is:

SB

SBNSB q

w , (35)

Net work of the SB steam cycle per kmol natural gas is then: SBNSBSBN wmW ,, kJ/kmoln.g (36)

where the mass of steam generated in the SB (mSB) is obtained from Eq. (3). The total net output of the combined cycle per kmol natural gas is: SBNHRBNGCNcom WWWW ,,, kJ/kmoln.g (37) The combined cycle thermal efficiency is then calculated as:

gngnP

comcom TCLHV

W

.., (38)

Another important parameter for the combined cycle is the power ratio, and it is defined as:

GCN

SBNHRBN

WWW

PR,

,, (39)

The output power produced by each combined cycle in (kW) can be determined from the following equation: gngncomcom MmWP .. kW (40) Table 2 Steam conditions of the cycle.

Parameter ST (SB) HP ST LP ST FWH1 FWH2 FWH3 Pressure (bar) Temperature (oC)

170 540

50 540

4.5 Tsat

59 -

14 -

1.9 -

4. Exergy Analysis

The exergy destruction in the different control volumes of the cycle is calculated by applying the exergy balance equation derived by [16, 17, 18] . This eauqtion reads:

CVCVooii WQ

TT

ememEXD 01 kW (41)

where,

CVQ : heat transferred to the control volume, kW

CVW : rate of work out from the control volume, kW T : temperature at which heat is transferred, K T0: reference temperature and equal to 298K.

The exergy of a flow stream for a given pressure (P) and temperature (T) is given by: ooo ssThhe (42) where, the properties h and s for steam are obtained from the present code, and for gas are calculated from the ideal gas model as:

oPo TTChh (43)

and

ooPo P

PRTTCss lnln (44)

The second-law efficiency for each control volume in steady state steady flow (SSSF) process is calculated as:

oind ExEx

EXD12 (45)

5. Results and Discussions

5.1 Effect of excess air factor

Figures 6 to 9 show the effect of three excess air factors (1.4, 1.6, and 1.8) on the performance and exergy destruction of the cycle, while the turbine inlet temperature is fixed (1300ºC). The thermal performance of the cycle is plotted against the compressor pressure ratio in Fig.6. The figure shows a noticeable effect of the excess air factor on the cycle performance. The output power and the power ratio decrease as the excess air factor increases, while the combined cycle efficiency increases with the increase of the excess air factor. As can be seen, an optimum compressor pressure ratio for the combined cycle efficiency is found depending on the value of the excess air factor. On the other hand, the change of the output power with the compressor pressure ratio is almost small. Figure 7 shows the exergy destructions in the cycle components at different excess air factors. The turbine inlet temperature and the compressor pressure ratio are fixed at 1300ºC and 15, respectively. It is obvious from the figure that the exergy destruction in the SB is the major part followed by that in the HRB. The SB destroys from 17% to 29% of the total exergy input while the HRB destroys from 3.5% to 13%. The exergy destruction in the SB decreases by increasing the excess air factor. This is due to the decrease in the amount of both fuel and steam in the SB. Also, by increasing the excess air factor, the exergy destruction in the HRB is slightly decreased due to the reduction in the temperature difference between hot (gas) and cold (steam) streams in the HRB. It is clear that the exergy destruction in the compressor is not affected by the excess air factor as the air mass flow rate is fixed constant. The exergy destructions in the ST, FWHs, and CON2 are decreased by increasing the excess air factor due to the decrease in the amount of steam generated in the SB, while those for the other components are almost unchanged.

30000

35000

40000

45000

50000

55000

60000

65000

70000

3 6 9 12 15 18 21 24 27 30 330.8

1

1.2

1.4

1.6

1.8

2

2.2

2.4

2.6

l1.4

l1.6

l1.8

πc

T3=1300ºC

PR

Pcom

PR

Pcom (kW)

(a) Output power and power ratio.

(b) Combined cycle efficiency.

Fig. 6 Thermal performance of the SBCC at different excess air factors.

Figure 8 shows the total exergy destruction in the combined cycle versus the compressor pressure ratio at the three selected excess air factors. Insignificant effect of the compressor pressure ratio on the total exergy destruction is found as shown in the figure. However, the excess air factor has significant effect. The total exergy destruction in the cycle decreases as the excess air factor increases. The total exergy destruction to the total exergy input of the combined cycle is found to be from 31% to 43%.

0.45

0.46

0.47

0.48

0.49

0.5

0.51

3 6 9 12 15 18 21 24 27 30 33

l1.4

l1.6

l1.8

πc

T3=1300ºC

ηcom

Fig. 7 Exergy destructions in the cycle components at different excess air factors.

Fig. 8 Total exergy destruction in the SBCC at different excess air factors.

Figure 9 shows a plot of the second-law efficiency with the compressor pressure ratio at different excess air factors. As shown in the figure, the second-law efficiency is increased by increasing the excess air factor. Also, the effect of the compressor pressure ratio is clear on the second-law efficiency. An optimum compressor pressure ratio is found depending on the value of the excess air factor.

0

5000

10000

15000

20000

25000

30000

35000

40000

C GT SB ST FWHs HRB HPST LPST CON1 CON2

λ=1.4

λ=1.6

λ=1.8

20000

25000

30000

35000

40000

45000

50000

55000

60000

3 6 9 12 15 18 21 24 27 30 33

l1.4

l1.6

l1.8

πc

T3=1300ºC

ExDcom (kW)

πc=15

T3=1300ºC ExDi

(kW)

Fig. 9 Second-law efficiency of the SBCC at different excess air factors.

5.2 Effect of turbine inlet temperature

Figures 10 to 13 show the effect of the turbine inlet temperature (1200ºC, 1300ºC, and 1400ºC) on the performance and exergy destruction of the cycle, while the excess air factor is fixed (1.6). Figure 10 shows the thermal performance of the cycle at the three different turbine inlet temperatures. The turbine inlet temperature has a high influence on the cycle performance as can be seen from this figure. Both the output power and the combined cycle efficiency are significantly increased by increasing the turbine inlet temperature, while the power ratio is decreased. The exergy destructions in the cycle components at different turbine inlet temperatures are shown in Fig. 11. For the GT, the exergy destruction decreases as the turbine inlet temperature increases. This is due to the reduction in the entropy change due to the decrease in the ratio (T3/T4). It is found that the exergy destruction in the SB is decreased by increasing the turbine inlet temperature. This is attributed to the decrease in the amount of steam generated in the SB. Also, the increase in the turbine inlet temperature increases the HRB exergy destruction. This is because the increase in the turbine inlet temperature increases the turbine exhaust temperature and the temperature difference between the hot and cold streams becomes larger which increases the HRB exergy destruction. By increasing the turbine inlet temperature, the amount of steam generated in the SB steam cycle is decreased and that of the HRB steam cycle is increased. This is why the exergy destructions in the SB steam cycle components (ST, FWHs, and CON2) are decreased while those in the HRB steam cycle (HPST, LPST, and CON1) are increased. The total exergy destruction in the combined cycle versus the compressor pressure ratio and at different turbine inlet temperatures is presented in Fig. 12. This figure shows a valuable influence of the turbine inlet temperature on the total exergy destruction. It is found that the total exergy destruction is noticeably decreased by increasing the turbine inlet temperature.

0.52

0.53

0.54

0.55

0.56

0.57

0.58

0.59

3 6 9 12 15 18 21 24 27 30 33

l1.4

l1.6

l1.8

πc

T3=1300ºC

η2nd

(a) Output power and power ratio.

(b) Combined cycle efficiency.

Fig. 10 Thermal performance of the SBCC at different turbine inlet temperatures.

Figure 13 present the effect of the turbine inlet temperature on the second-law efficiency of the combined cycle. As can be seen, the second-law efficiency is highly affected by the turbine inlet temperature. The second-law efficiency is significantly increased by increasing the turbine inlet temperature.

40000

45000

50000

55000

60000

65000

3 6 9 12 15 18 21 24 27 30 330.6

0.8

1

1.2

1.4

1.6

1.8

2

2.2

2.4

T3=1200ºC

T3=1300ºC

T3=1400ºC

0.44

0.45

0.46

0.47

0.48

0.49

0.5

0.51

0.52

3 6 9 12 15 18 21 24 27 30 33

T3=1200ºC

T3=1300ºC

T3=1400ºC

πc

λ =1.6

πc

λ =1.6 Pcom

PR

Pcom (kW)

PR

ηcom

Fig. 11 Exergy destructions in the cycle components at different turbine inlet temperatures.

Fig. 12 Total exergy destruction in the SBCC at different turbine inlet temperatures.

A comparison between the SBCC and the conventional combined cycle was carried out. This comparison aims to evaluate the performance of this combined cycle. Furthermore, this comparison was performed at a fixed air mass flow rate of 67.9268 kg/s, turbine inlet temperature of 1300ºC, and the other parameters are as listed in Table 1.

0

5000

10000

15000

20000

25000

30000

35000

C GT SB ST FWHs HRB HPST LPST CON1 CON2

T3=1200ºC

T3=1300ºC

T3=1400ºC

35000

37000

39000

41000

43000

45000

47000

49000

51000

53000

55000

3 6 9 12 15 18 21 24 27 30 33

T3=1200ºC

T3=1300ºC

T3=1400ºC

πc

λ =1.6

ExDcom (kW)

πc=15 =1.6 ExDi

(kW)

Fig. 13 Second-law efficiency of the SBCC at different turbine inlet temperatures.

Figure 14 shows the total output power for the two combined cycles versus the compressor pressure ratio at a fixed turbine inlet temperature. This figure shows that the SBCC has higher values of the output power compared with the conventional combined cycle. The values of the output power are found to be from 1.6 to 2.1 times of those of the conventional combined cycle as shown in Fig. 12. It is also found that higher values of the output power for the SBCC are predicted at the lower excess air factor.

Fig. 14 Comparison between the output power of the SBCC and the conventional cycle.

0.51

0.52

0.53

0.54

0.55

0.56

0.57

0.58

0.59

0.6

0.61

3 6 9 12 15 18 21 24 27 30 33

T3=1200ºC

T3=1300ºC

T3=1400ºC

0

10000

20000

30000

40000

50000

60000

70000

80000

90000

3 6 9 12 15 18 21 24 27 30 33

Conventional

λ=1.2 SBCC

λ=2.2 SBCC

πc

λ =1.6

η2nd

πc

T3=1300ºC

Pcom (kW)

Figure 15 shows a plot of the combined cycle efficiency with the compressor pressure ratio for the two combined cycles. As can be seen, the combined cycle efficiency of the SBCC is lower than that of the conventional combined cycle. Also, optimum compressor

pressure ratios that give maximum efficiencies are predicted for the SBCC. However for the conventional combined cycle, the efficiency is continuously increased by increasing the compressor pressure ratio.

Fig. 15 Comparison between the combined cycle efficiency of the SBCC and the conventional cycle.

Figure 16 shows a comparison between the total exergy destruction in the two combined cycles. It is found that the minimum values of the exergy destruction in the SBCC are found at the higher excess air factor (λ=2.2). At this condition, the total exergy destruction in the SBCC is lower than that in the conventional combined cycle for the whole range of the investigated compressor pressure ratio. The results reveal that the reduction in the total exergy destruction in the SBCC over the conventional combined cycle reaches up to 33%. Figure 17 shows a comparison between the second-law efficiency of the SBCC and that of the conventional combined cycle. The second-law efficiency of the SBCC is seen to be higher than that of the conventional combined cycle when the excess air factor is equal to 2.2. This means that the SBCC has the potential to give a larger increase in the second-law efficiency. The results also show that there is an improvement in the second-law efficiency of the SBCC over the conventional combined cycle of about from 9.5% to 18.5%.

0.4

0.42

0.44

0.46

0.48

0.5

0.52

0.54

0.56

3 6 9 12 15 18 21 24 27 30 33

Conventional

λ=1.2 SBCC

λ=2.2 SBCC

πc

T3=1300ºC

ηcom

Fig. 16 Comparison between the total exergy destruction in the SBCC and the conventional cycle.

Fig. 17 Comparison between the second-law efficiency of the SBCC and the conventional cycle.

Finally, the present predictions for the SBCC are correlated in terms of the investigated parameters. These correlations are performed for the combined cycle efficiency, second-law efficiency, and the total exergy destruction ratio (the total exergy destruction to the total exergy input) as follows: 321

030aaa

C TTa (46)

0

10000

20000

30000

40000

50000

60000

70000

80000

3 6 9 12 15 18 21 24 27 30 33

Conventional

λ=1.2 SBCC

λ=2.2 SBCC

0.45

0.47

0.49

0.51

0.53

0.55

0.57

0.59

0.61

0.63

3 6 9 12 15 18 21 24 27 30 33

Conventional

λ=1.2 SBCC

λ=2.2 SBCC

πc

T3=1300ºC

ExDcom (kW)

πc

T3=1300ºC

η2nd

where the variable is one of ηcom , η2nd ,or *

comEXD and the coefficients ,,, 210 aaa and

3a are listed in Table 3. T3 and T0 are temperatures in (K). The above correlation is valid for ranges of the operating parameters of 6 πc 30, 1200ºC T3 1400ºC, and 1.2 λ 2.0.

Table 3 Coefficients of the correlation 46.

% DEVmax 3a 2a 1a 0a Variable

2.57 0.62025 0.13231 1.137E-2 0.15798 com

2.76 0.71240 0.18693 1.698E-2 0.15204 nd2

4.75 -1.04413 -0.30428 -2.806E-2 2.60315 *comEXD

6. Conclusions

A thermodynamic analysis of a supercharged boiler combined cycle (SBCC) was carried out. The effects of the inlet temperature of the gas turbine, the excess air factor, and the compressor pressure ratio on the performance of the cycle were investigated. A comparison study among the SBCC and the conventional cycle leads to the following conclusions:

1. The largest values of the output power for the SBCC are predicted at a minimum excess air factor and a maximum turbine inlet temperature.

2. The SBCC has higher values of the output power compared with the conventional combined cycle, and these values are found to be from 1.6 to 2.1 times of the conventional cycle.

3. The values of the combined cycle efficiency of the SBCC are lower than that of the conventional cycle.

4. For a certain value of the turbine inlet temperature, in the SBCC, optimum compressor pressure ratios which give maximum efficiencies are predicted. While, for the conventional cycle, the efficiency is continuously increased with the compressor pressure ratio.

5. The exergy analysis of various components for all studied combined cycle shows the maximum exergy losses to be in the supercharged boiler and the heat recovery boiler, thus the need to make efforts to minimize losses in these components.

6. Lower values of the total exergy destruction in the SBCC are observed at the higher excess air factor (λ=2.2).

7. The total exergy destruction is found to be lesser in the SBCC than in the conventional combined cycle. On the other hand, the exergy destruction ratio of the SBCC is found from 31% to 43%, while that for the conventional cycle is from 43% to 52% of the exergy input.

8. The SBCC has higher values of the second-law efficiency than the conventional cycle. The improvement in the second-law efficiency of the SBCC is found from 9.5% to 18.5% compared with the conventional cycle.

9. Finally, the turbine inlet temperature, the excess air factor, and the compressor pressure ratio strongly influence the combined cycle performance.

REFERENCES 1- Briesch, M. S., Bannister, R. L., Diakunchak, I. S., and Huber, D. J., A Combined

Cycle Designed to Achieve Greater Than 60 Percent Efficiency, ASME J. Engineering for Gas Turbines and Power, Vol. 117, pp. 734-741, 1995.

2- Rice, I. G., The Combined Reheat Gas Turbine/Steam Turbine Cycle. Part I-A Critical Analysis of the Combined Reheat Gas Turbine/Steam Turbine Cycle, ASME J. Engineering for Power, Vol. 102, pp. 35-41, 1980.

3- Rice, I. G., The Combined Reheat Gas Turbine/Steam Turbine Cycle. Part II-The LM 5000 Gas Generator Applied to the Combined Reheat Gas Turbine/Steam Turbine Cycle, ASME J. Engineering for Power, Vol. 102, pp. 42-49, 1980.

4- Rice, I. G., The Reheat Gas Turbine With Steam-Blade Cooling-A Means of Increasing Reheat Pressure, Output, and Combined Cycle Efficiency, ASME J. Engineering for Power, Vol. 104, pp. 9-22, 1982.

5- Bassily, A. M., Modeling, Numerical Optimization, and Irreversibility Reduction of a Dual-Pressure Reheat Combined-Cycle, Applied Energy, Vol. 81, pp. 127-151, 2005.

6- Akiba, M., and Thani, E. A., Thermodynamic Analysis of New Combination of Supercharged Boiler Cycle and Heat Recovery Cycle for Power Generation, ASME J. Engineering for Gas Turbines and Power, Vol. 118, pp. 453-460, 1996.

7- Mikhael, N. N., Morad, K. K. A., and Mohamed, A. M. I., Design Criterion of Solar-Assisted Combined-Cycle Power Plants with Parabolic Through Concentrators, Port-Said Engineering Research J., Vol. 4, pp. 80-101, 2000.

8- Ghazikhani, M., Takdehghan, H., and Moosavi, A., Exergy Analysis of Gas Turbine Air-Bottoming Combined Cycle for Different Environment Air Temperature, Proceedings of 3rd International Energy, Exergy and Environment Symposium, 2007.

9- Koch, C., Cziesla, F., and Tsatsaronis, G., Optimization of Combined Cycle Power Plants Using Evolutionary Algorithms, Chemical Engineering and Processing, 2007.

10- El-Masri, M. A., and Magnusson, J. H., Thermodynamics of Isothermal Gas Turbine Combined Cycle, ASME J. Engineering for Gas Turbines and Power, Vol. 106, pp. 743-749, 1984.

11- Jericha, H., and Hoeller, F., Combined Cycle Enhancement, ASME J. Engineering for Gas Turbines and Power, Vol. 113, pp. 198-202, 1991.

12- Bolland, O., A Comparative Evaluation of Advanced Combined Cycle Alternatives, ASME J. Engineering for Gas Turbines and Power, Vol. 113, pp. 190-197, 1991.

13- Seyedan, B., Dhar, P. L., Guar, R. R., and Bindra, G. S., Optimization of Waste Heat Recovery Boiler of a Combined Cycle Power Plant, ASME J. Engineering for Gas Turbines and Power, Vol. 118, pp. 561-564, 1996.

14- Sanjay, Y., Singh, O., and Prasad, B. N., Energy and Exergy Analysis of Steam Cooled Reheat Gas–Steam Combined Cycle, Applied Thermal Engineering, 2007.

15- El-Wakil, M. M., Power Plant Technology, McGraw-Hill, Inc., 2nd ed., 1988. 16-Wylen, G. V., Sonntag, R., and Borgnakke, C., Fundamentals of Classical

Thermodynamics, John Wiley & Sons, Inc., 4th ed., 1994. 17- Ebadi M.J, Gorji-Bandpy M., Exergetic Analysis of Gas Turbine Plants, Int. Journal of

Exergy Research.Vol. 2(1): pp. 31-39, 2005. 18- Ayoola1, P. O and Anozie, N.A, A Study of Sections Interaction Effects on

Thermodynamic Efficiencies of a Thermal Power Plant, British Journal of Applied Science & Technology, ISSN: 2231-0843, Vol.3(4): pp. 1201-1214, 2013.