7. Brayton Cycle and Combined Cycle - Engsoft · 2018-01-13 · Thermodynamics 7. Brayton Cycle 4 /...

Thermodynamics 7. Brayton Cycle 1 / 126

7. Brayton Cycle and Combined Cycle


Starter &

Gear Box

Air Inlet CompressorCombustor

Turbine Exhaust

VIGV

Air Extraction

PortsDiffuser

Transition

Piece

Cold Section Hot Section

7FA Gas Turbine for Power Generation


Idealized Brayton Cycle [1/3]


The entering air is compressed to higher pressure.

No heat is added. However, compression raises the air temperature so that the discharged air has

higher temperature and pressure.

The mechanical energy transmitted from the turbine is used to compress the air.

Compression Process (1 2)

Compressed air enters the combustor, where fuel is injected and combustion occurs.

Combustion occurs at constant pressure. However, pressure decreases slightly in the practical process.

Although high local temperatures are reached within the primary combustion zone (approaching

stoichiometric conditions), the combustion system is designed to provide mixing, burning, dilution,

cooling.

Combustion mixture leaves with mixed average temperature.

The chemical energy contained in the fuel is converted into thermal energy.

Combustion of fuel is irreversible process, and entropy is produced during the combustion of the fuel.

Combustion Process (2 3)



The thermal energy contained in the hot gases is converted into mechanical work in the turbine.

This conversion actually takes place in two steps:

• Nozzle: the hot gases are expanded and accelerated, and a portion of the pressure energy is

converted into kinetic energy.

• Bucket: a portion of the kinetic energy is transferred to the rotating buckets and converted into

mechanical work.

Some of the work produced by the turbine is used to drive the compressor, and the remainder is used to

drive load equipment, such as generator, ship propeller, and pump, etc.

Typically, more than 50% of the work produced by the turbine section is used to power the compressor.

Expansion Process (3 4)

Exhaust Process (4 1)

This is a constant-pressure cooling process.

This cooling is done by the atmosphere, which provides fresh, cool air as well.

The actual cycle is an “open” rather than “closed”.



Simple Cycle

The term simple cycle is used to distinguish this configuration from the complex cycles, which utilizes

additional components, such as heat exchanger for regeneration, intercooler, reheating system, or steam

boilers.

This cycle is suitable for a fixed speed and fixed load operation, such as power generation.

In order to analyze gas turbine system in a convenient form, the assumptions listed below are frequently

used.

1) The working fluid is treated as the air. The air is an ideal gas and has a constant specific heat. (In

practice, there is a change in the composition of the working fluid because of the combustion process)

2) The combustion process is replaced by a heat transfer process from an external source. In other words,

the mass flow rate remains constant throughout the system.

3) The inlet and exhaust processes are replaced by a constant pressure process that will in turn complete

the gas turbine cycle.

4) All processes are internally reversible.

The combination of these assumption is called the air-standard cycle approach.

General Notes


Heat and Work

754 MJ/s

(100%)

205 MW

(27.2%)

203 160 119 MW = 482 MW (63.9%)

277 MW (Net Output)

(36.7%)

272 MJ/s

(36.1%)

Simple Cycle


Cycle Arrangement

Simple Cycle

Compressor

Fuel Combustor

Turbine

Air

Power

Exhaust gas1

243

p

2

1

T

(h)

s

qin

3

41

2

3

4

qout

winwout

win

wout

qin

qout


Compressor Work (|CW|) = Turbine Work (|TW|)

Coo hhCW /1,2,

Coop TTc /1,2,

11

1

1,

2,1,

1,

2,1,

o

o

C

op

o

o

C

op

p

pTc

T

TTc

1

1

1,

CPR

Tc

C

op

4,3, ooT hhTW

4,3, oopT TTc

1

3,

4,

3,

3,

4,

3, 11o

o

opT

o

o

opTp

pTc

T

TTc

1

3, 1 TPRTc opT

11

1

3,

1,

1

CPR

T

TTPR

oTC

o

h

s

1

2

3

4

3

4

2

Compressor – Turbine Matching

Compressor

Fuel Combustor

Turbine

Air

Power

Exhaust gas1

24

3


Process Component Heat Work Process

12 Compressor q12 = qC = 0 w12 = wC = (h2h1) Power in (adiabatic compression)

23 Combustor q23 = qB = h3h2 w23 = wB = 0 Heat addition at constant pressure

34 Turbine q34 = qT = 0 w34 = wT = h3h4 Power out (adiabatic expansion)

41 Exhaust q41 = qE = (h4h1) w41 = wE = 0 Heat release at constant pressure

Simple Cycle Analysis [1/13]

121212 whhq

p

2

1

T

(h)

s

qin

3

41

2

3

4

qout

winwout

win

wout

qin

qout


0 5 10 15 20Pressure Ratio [r]

0.0

0.1

0.2

0.3

0.4

0.5

0.6

0.7

0.8

Th

erm

alE

ffic

ien

cy

Thermal efficiency in a simple cycle gas

turbine increases with pressure ratio

and specific heat ratio.

The increasing rate of the thermal

efficiency is getting smaller as the

pressure ratio increases.

4

3

1

2

p

p

p

pr

crth

111

/1

1

rc

Thermal Efficiency


2

1

23

14

23

1423

23

4123

23

3412 11T

T

TT

TT

hh

hhhh

q

qq

q

ww

q

w

inputheat

ouputworknet

in

sys

th


Specific Work Output


1

3

T

Tt

Specific work output:

124341342312 TTcTTcwwwww ppsys

1

111

11 /1

/1

1

c

ctr

rt

Tc

w

p

sys

The specific work output, which is the output per unit mass flow rate of working fluid, is a function of

pressure ratio and maximum cycle temperature.

The specific work output increases with the pressure ratio when the maximum cycle temperature is greater

than a certain value.

There is a pressure ratio having a maximum specific work out in a constant t-curve.

4

3

1

2

p

p

p

pr

1

rc


0.0

0.5

1.0

1.5

2.0

Sp

ecific

Wo

rkO

utp

ut[w

/CT

]sys

1p

t = 2

t = 3

t = 4

t = 5



Optimum Pressure Ratio for a Given TIT [1/3]

There are different optimum pressure ratios in

terms of thermal efficiency and specific work

output for a given maximum cycle temperature.

The thermal efficiency increases with the

pressure ratio, and it has a maximum value

when the air temperature at the compressor

outlet is equal to TIT.

In this limiting case, the cycle net work tends

toward zero, and the thermal efficiency

approaches the Carnot efficiency. In terms of

available work output (= turbine work –

compressor work), it increases with pressure

ratio and reaches a maximum at a certain

pressure ratio, then it becomes smaller, and

finally reaches zero when the air temperature at

the compressor outlet is equal to TIT.

Therefore, it is clear that there are different optimum pressure ratios in terms of thermal efficiency and

specific work output for a given maximum cycle temperature.

This means that the maximum net specific work and the maximum thermal efficiency do not occur at the

same pressure ratio. Therefore, in designing gas turbines, the design pressure ratio must be a compromise

between the maximum thermal efficiency and the maximum specific work.

T

(h)

s

1

2

3

4

4

3

2

2

4

3

Tmax



This can be understand more easily with the

investigation of the diagram shown relationship among

thermal efficiency, specific output, and pressure ratio.

In a real (non-ideal) gas turbine, the specific work

produced by the turbine increases more rapidly than

the specific work that the compressor consumes, as

the pressure ratio increases (from 5 to 8).

As a result, both shaft specific work output and

efficiency increase with pressure ratio.

As the pressure ratio increases, the compressor outlet

temperature also increases. This in turn reduces the

combustor temperature rise necessary to achieve a

given TIT. As a result, the combustor heat input

decreases and cycle efficiency increases.


Specific output, MW/lb/s

Th

erm

al e

ffic

ien

cy

Maximum thermal efficiency

Maximum shaft output

Pressure ratio = 5

1015

20

TIT = 1,800F (982C)



At the point corresponding to the maximum shaft specific work output (turbine specific work produced

minus compressor specific work consumed), the turbine specific work produced and the compressor

specific work consumed increase at the same rate.

As pressure ratio increases further (beyond the maximum shaft output), the compressor specific work

consumed increases as a greater rate than the turbine specific work produced and, as a result, the shaft

work output actually decreases.

However, because the combustor heat input continues to decrease, the efficiency continue to increase as

pressure increases (from 8 to 16). The reason for this is that the shaft specific work output decreases

because compressor specific work consumed is increasing faster than turbine specific work output, as the

pressure ratio increases. Therefore, shaft specific work output decreases.

However, fuel consumption is reduced and overall thermal efficiency increases because of higher

compressor outlet temperatures due to higher pressure ratios.

As the pressure ratio increases further (greater than 16), the increase in compressor specific work

consumed offsets the advantage of higher compressor outlet temperature and overall efficiency begins to

decrease.

The actual pressure ratio at which this occurs depend on the specific gas turbine considered.



The Performance Map of a Simple-Cycle Gas Turbine



Practical Brayton Cycle

h1 = enthalpy at compressor inlet

h2s = enthalpy at constant entropy and

compressor discharge pressure

h2 = actual enthalpy at compressor discharge

pressure

h3 = enthalpy at turbine inlet

h4s = enthalpy at constant entropy and turbine

exit pressure

h4 = actual enthalpy at turbine exit pressure

p

2

1

T

(h)

s

qin

3

4 1

2s

3

4

qout

win

wout

2

4s

12

12

hh

hh sC

s

Thh

hh

43

43



General Expression of the Simple Cycle Net Work and Efficiency [1/4]

1

1

1

21

12T

TTcTTcw s

C

p

sp

C

C

crp

p

T

T s

1

1

1

2

1

2

where, r is a pressure ratio. Therefore, compressor

work is

11 c

Tcw

C

p

C

Similarly, the turbine work is given by

cTcTTcw pTspTT

11343


T

(h)

s

1

2s

3

4

4s

2



CTnet www

11

1

c

tc

Tcw TC

C

p

net

1

2

1

3123

T

T

T

TTcTTcq ppin

12

12

TT

TT sC

1

11

1

2 cT

T

C

where, t is a maximum temperature ratio. Therefore, heat input is expressed by

111

23 ctTc

TTcq C

C

p

pin

11

1

ctc

ctc

q

w

C

TC

in

netcy

Four major parameters affecting the cycle efficiency of a simple gas turbine are pressure ratio (c), TIT

(t), compressor efficiency, and turbine efficiency.

1

3

T

Tt




Pressure Ratio

t = 3.25

t = 4.73

t = 5.69

Reversible

0.7

0.1

Th

erm

al E

ffic

ien

cy

1

3

T

Tt

2 4 6 80 10 12 14 16 18 20 22 24 26 28 30

0.6

0.5

0.4

0.3

0.2

0

T = 0.90

C = 0.85




In general, the cycle efficiency is relatively low, because of the high EGT, and because a significant portion

of the turbine output is used for compressor operation.

For a given turbine and compressor efficiency, the cycle performance is determined by the TIT and pressure

ratio.

The TIT is usually fixed by the metallurgical temperature limit of the first row of turbine blade.

As the TIT increases, the cycle efficiency is greatly improved.

The impact of the pressure ratio on the cycle efficiency is quite different.

There is an optimum pressure ratio that produces the maximum cycle efficiency.

The optimum pressure ratio increases with the TIT.



Simple Cycle Efficiency

Shaft power application



Simple Cycle

[Exercise 4.1]

Calculate the cycle efficiency and net work per pound of air. A gas turbine is operated under the

following conditions.

• Compressor inlet pressure and temperature 14.7 psia, 60F

• Pressure ratio 12

• Compressor efficiency 0.88

• Turbine inlet temperature 2000F

• Turbine efficiency 0.90

• Average constant pressure specific heat 0.25 Btu/lb-R

• Specific heat ratio 1.4

The pressure drops in the combustor, compressor inlet, and turbine outlet are assumed to be negligible.

T

(h)

s

1

2s

3

4

4s

2


[Solution]

Process 12s is an isentropic process. Thus,

T2s = T1 2.03394 = 1057.6 R

T2 = 1130.9 R

Similarly, process 34s is an isentropic process. Thus,

T4s = 1209.5 R

T4 = 1334.6 R

03394.212 4.1

4.01

1

2

1

2

p

p

T

T s

12

12

TT

TT sC

1

4

3

4

3

ss p

p

T

T

s

TTT

TT

43

43

Simple Cycle


[Solution]

Heat input in the combustor is

Compressor work and turbine work are

Cycle efficiency is

cy = 0.387 or 38.7%

Cycle network is

128.6 Btu/lb

23 TTcq pin

12 TTcw pC

43 TTcw pT

23

141TT

TT

q

ww

in

CTcy

CTnet www

11

1

ctc

ctc

q

w

C

TC

in

netcy

Simple Cycle


Compressor

Fuel

Combustor

Turbine

Air

Power

Exhaust gas

Heat exchanger

1

2

5

34

6

Regenerative Cycle

Cycle Arrangement

p

2

1

T

s

3

41

2

3

46

5

5

6

A B C D


General Notes

Regenerative Cycle

Normally, gas turbine exhaust gas temperature is higher than that of air discharging from a compressor.

In order to increase the gas turbine efficiency, the gas turbine exhaust gas can be used to heat the air leaving

the compressor, thus reducing the amount of fuel required to reach the firing temperature.

This is achieved by the use of regenerators or recuperators, which heat the compressor exit air by the

exhaust gases from the turbine exit.

Regenerators or recuperators are usually used in small- to intermediate-sized gas turbines having output less

than 10 MW.

Regenerative cycle is also called as heat exchange cycle.


A Typical Regenerative Cycle Gas Turbine

Air flow path – Mercury 50 gas turbine (Solar)

Power = 4.6 MW

PR = 9.9:1

TIT = 2200F (1204C)

th = 38.5%

Regenerative Cycle


Regenerative Cycle Analysis [1/5]

Cycle Analysis



25 Heat Ex. q52 = qin = h5h2 w52 = wHE = 0 Heat addition through heat exchanger

53 Combustor q53 = qB = h5h3 w53 = wB = 0 Heat addition in a burner


46 Heat Ex. q46 = qout = (h6h4) w46 = wHE = 0Heat transfer to compressor discharged air

(h4h6 = h5h2)

61 Exhaust q61 = qE = (h6h1) w61 = wE = 0 Heat release to atmosphere

121212 whhq

p

2

1

T

s

3

41

2

3

46

5

5

6

A B C D



0.0

0.5

1.0

1.5

2.0S

pecific

Wo

rkO

utp

ut[w

/CT

]sys

1p

t = 2

t = 3

t = 4

t = 5

압력비 vs 비출력

The regenerative cycle has

the exactly same specific

work out with a simple cycle.

11

11

c

ct

Tc

w

p

sys

1

3

T

Tt

1

rc

1

2

p

pr



압력비 vs 열효율

t

cth 1

1

3

T

Tt

1

rc1

2

p

pr

4

1

3

2 11T

T

T

Tth



0

0.1

0.2

0.3

0.4

0.5

0.6

0.7

0.8

0.9

Th

erm

alE

ffic

ien

cy

Simple cycle efficiency

t = 2

t = 3

t = 4

t = 5t = 6

t

cth 1

사이클 최고온도가 일정한 상태에서압력비가 작아질수록 재생사이클 가스터빈 열효율 향상.

압력비가 일정한 상태에서 최고온도가 증가할수록 열효율이 향상된다.

4

1

3

2 11T

T

T

Tth

재생사이클 가스터빈은 TIT가 증가할수록, 그리고 EGT가 증가할수록 열효율 향상.


The cycle efficiency decreases as the pressure ratio increases, which is opposite to that of a simple cycle.

This is due to the fact that, as the pressure ratio increases the air temperature exiting the compressor

increases and ultimately will exceed that of the turbine exhaust gas temperature. Then heat in the heat

exchanger (regenerator) will be lost from the air to the exhaust gases instead of desired gain.

The efficiency with lower temperatures, say at t=2, is seen to become negative soon after the pressure

ratio 11.3 is exceeded. The reason is that the temperature at compressor outlet actually exceeds the

assumed combustion temperature in this case.

In many cases, regeneration is not desirable. This is because the efficiency increases with pressure ratio, if

regenerative cycle is not employed.

Efficiency, with regenerative cycle rises very rapidly with increase in maximum temperature of the cycle.

Lower pressure ratios and high cycle temperatures are favorable for the regenerative cycle, since a large

heat recovery is then possible.

After the efficiency becomes equal to that of simple cycle, any further increase of pressure ratio will yield an

efficiency which is lower than that of simple cycle and that is of no interest.

Power output may be reduced by 10% for a given size of plant because of the pressure losses occurred in

the heat exchanger.

Normally, micro gas turbines having lower thermal efficiency adopts this kind of arrangement to improve

thermal efficiency.



Cycle Efficiency of Practical Regenerative Cycles


24

25

24

25

TT

TT

hh

hhreg

53

1243

TT

TTTT

q

ww

in

CTcy

2423

1243

TTTT

TTTT

reg

cy

1

2

1

3

3

4

1

2

1

3

1

2

1

3

3

4

1

3 1

T

T

T

T

T

T

T

T

T

T

T

T

T

T

T

T

T

T

reg

cy

11

1

2

C

c

T

T

cT

TT

111

3

4

It can be seen that the irreversibility of the system significantly lower the cycle efficiency.

Compared with the simple gas turbine system, the optimum pressure ratio is smaller for the regenerative

cycle.

The small pressure ratio means a small cycle net output.

Therefore, the cost associated with this output reduction must be weighted against the saving that can be

affected by the cycle efficiency improvement.

Pressure Ratio

t = 4.73

t = 5.69

Reversible0.7

0.1

Therm

al E

ffic

iency

2 4 6 80 10 12 14 16 18 20 22 24 26 28 30

0.6

0.5

0.4

0.3

0.2

0

1

3

T

Tt

T = 0.90

C = 0.85

reg = 0.80

0.8

0.9

1.0

Practicalt = 5.69

t = 4.73


Regenerative Cycle

This modified form is more suitable for fuels those combustion products contain constituents which may

corrode or erode the turbine blades.

It is much less efficient than the simple cycle power plant because of the efficiency of the heat exchanger.

Such a cycle may be considered only if low grade fuels are to be used.

Gas turbines of a more complicated design (i.e., with intermediate cooling in the compressor or

recuperator) are less suitable for combined cycle plants. They normally have a high simple cycle

efficiency combined with a low exhaust gas temperature, so that the efficiency of the water/steam cycle

is accordingly lower.

An Alternative Regenerative Cycle

Compressor

FuelCombustor

Turbine

Air

Power

Exhaust gasHeat exchanger

1

2 34


Reheat Cycle

Cycle Arrangement

FuelReheat combustor

LP turbine

Power

Exhaust gas5

4 6

Compressor

Fuel

Combustor

Air

1

2

3

HP turbine

p

2

1

T

s

3

61

2

5

6

4 54

3

4

4


A Typical Reheat Cycle Gas Turbine for Power Generation - GT26 Gas Turbine

Reheat Cycle


Reheat Cycle

F-14

Reheat Cycle Gas Turbine for Military Aviation

F100-PW-229 (F-15, F-16)Reheat Cycle Gas Turbine for Civil Aviation

Olympus 593


Cycle Analysis

Reheat Cycle Analysis [1/3]




34 Turbine q34 = qHPT = 0 w34 = wHPT = h3h4 Power out (adiabatic expansion)

45 Reheat q45 = qin,R = h5h4 w45 = wR = 0 Heat addition in a reheat combustor


121212 whhq

p

2

1

T

s

3

61

2

5

6

4 54

3

4

4



11

121

max

c

ct

TC

w

p 1

3

T

Tt

1

rc1

2

p

pr


0.0

0.5

1.0

1.5

2.0

2.5

Sp

ecific

Wo

rkO

utp

ut[w

/CT

]

t = 2

t = 3

t = 4

t = 5

Reheat cycle

Simple cycle

sys

p1


0.0

0.1

0.2

0.3

0.4

0.5

0.6

Th

erm

alE

ffic

ien

cy

Simple cycleefficiency

t = 2

t = 3

t = 4

t = 5

ctct

cctth

/2

1/112


An increase of specific work output can be obtained by splitting the expansion and reheating the gas

between the high pressure and low pressure turbines.

The increase of specific work output can be seen in p- diagram.

The turbine work increase is obvious from the fact that the vertical distance between any pair of constant

pressure lines increases as the entropy increases. Thus,

(T3T4) + (T5 T6) (T3 T4’)

The shaft length becomes longer and the control of shaft vibration becomes difficult.

The maximum temperature in low pressure turbine is the same as in high pressure turbine.

Thermal efficiency of the reheat cycle is lower than that of the simple cycle. This is because the reheat

cycle is made by the combination of a simple cycle and a less efficient cycle which is operated over a lower

temperature range.



Intercooled Cycle

Cycle Arrangement

HP compressor

Fuel

Combustor

Power turbine

Air

Power

Intercooler

Coolant in Coolant out

HPT

LP compressor

LPT

1 2 3 4 5

6

p

2

1

T

s

3

6 1

2

5

6

4 5

4

3

2

2


Typical Intercooled Cycle Gas Turbines

Air-to-Water Heat Exchanger Air-to-Air Heat Exchanger

Intercooler

LMS 100 (GE)


The specific work output of a gas turbine may be improved substantially by reducing the work of

compression.

If, therefore, the compression process is carried out with intercooling, the work of compression will be

reduced, as can be seen in p- diagram.

The compressor work decrease is obvious from the fact that the vertical distance between any pair of

constant pressure lines decreases as the entropy decreases. Thus,

(T2T1) + (T4 T3) < (T2′ T1)

Heat is extracted by an intercooler between the first and second compressors.

Rejecting heat worsens SFC, since more fuel should be burnt to raise cooler compressor delivery air to any

given TIT.

Therefore, the thermal efficiency of the intercooled cycle will be less than that for a simple cycle.

Intercooling is useful when the pressure ratios are high and the efficiency of the compressor is low.

Intercooled Cycle


Intercooled Cycle Analysis [1/5]

Cycle Analysis


12 LP compressor q12 = qC = 0 w12 = wLPC = (h2h1) Power supply in a LP compressor

23 Intercooler q23 = qIC = (h3h2) w23 = wIC = 0 Heat rejection in an intercooler

34 HP compressor q34 = qC = 0 w34 = wHPC = (h4h3) Power supply in a HP compressor




121212 whhq

p

2

1

T

s

3

61

2

5

6

4 5

4

3

2

2


압력비 vs 비출력

Simple CycleIntercooled Cycle


0.0

0.5

1.0

1.5

2.0

Sp

ecific

Wo

rkO

utp

ut[w

/CT

]

t = 2

t = 3

t = 4

t = 5

max

p1


0.0

0.5

1.0

1.5

2.0

Sp

ecific

Wo

rkO

utp

ut[w

/CT

]sys

1p

t = 2

t = 3

t = 4

t = 5

121

max cc

tt

Tc

w

p 1

3

T

Tt

1

rc



압력비 vs 효율


0.0

0.1

0.2

0.3

0.4

0.5

0.6

Th

erm

alE

ffic

ien

cy

Simple cycleefficiency

t = 2

t = 3

t = 4

t = 5


0.0

0.1

0.2

0.3

0.4

0.5

0.6

Th

erm

alE

ffic

ien

cy

t = 2

t = 3

t = 4

t = 5

Intercooled Cycle Reheat Cycle

1

3

T

Tt

1

rcct

cctth

2/1



A Typical Spray Intercooling Gas Turbine

LM6000-SPRINT Gas Turbine

LM6000 (GE)



Spray Intercooling

Intercooling can also be accomplished by fog spraying atomized water between the HP and LP

compressors.

GE LM6000-SPRINT is one example of such a system.

Water in injected through 24 spray nozzles.

Water is atomized to a droplet diameter of less than 20 microns using high-pressure air taken from the

eighth-stage of HP compressor.

Injecting water significantly reduces the compressor outlet temperature.

The result is higher output and better efficiency.

Output increases of more than 20% and efficiency increases of 3.9% are possible on 90F(32C) day.



[ A simple schematic diagram of a combined cycle system ]

Compressor

Fuel Combustor

Turbine

Air

Steam Turbine

Exhaust gas1

2

4

3

G G

HR

SG

5

6

7

8

9Condenser

Pump

Combined Cycle

Cycle Arrangement


Combined cycle power plants have a higher

thermal efficiency because of the application of

two complementary thermodynamic cycles

Combined Cycle

Condenser

(heat out)

T

s

Topping cycle

Bottoming cycle

Combustion

(heat In)

Stack

(heat

out)


Condenser

G

G

Fuel

Air

Gas Turbine

Heat Recovery

Steam Generator

IP Steam LP

SteamCold Reheat

Steam

Hot Reheat

Steam

Main

Steam

Steam Turbine

Condensate Pump

Steam

Water

Fuel

Air

Three Pressure

Reheat Cycle

Combined Cycle


The first power generation gas turbine was introduced by ABB in 1937.

It was a standby unit with a thermal efficiency of 17%.

Today the gas turbine is a major player in the huge power generation market, with orders of around 30GW

per year.

This success is due partly to large reserves of natural gas which provide a cheap fuel which is rich in

hydrogen, and therefore produces less carbon dioxide than liquid fuels.

The other factor is thermal efficiency, which for combined cycle power plants have 60%

A final advantage is the variety of gas turbines in a very wide range of power levels, up to 300MW per simple

cycle, and 500MW per combined cycle.

Power Generation Applications


Unit Steam CycleNet Plant

Output (MW)

Heat Rate (LHV) Thermal

EfficiencyBtu/kWhr kJ/kWhr

S106B Non-Re, 3-P 59.8 7005 7390 48.7

S106FA Re, 3-P 107.1 6440 6795 53.0

S107EA Non-Re, 3-P 130.2 6800 7175 50.2

S107FA Re, 3-P 258.8 6090 6425 56.1

S107G Re, 3-P 350.0 5885 6210 58.0

S107H Re, 3-P 400.0 5690 6000 60.0

The three-pressure reheat cycle that is applied with gas turbines with exhaust gas

temperature higher than 1000F/538C.

= ---------------3412.14

Heat Rate3412.14 = constant for the number of BTU’s in a kilowatt

(1 kWh = 3412.14 Btu)

GE’s Single-Shaft STAG Ratings


Simple cycle

Combined cycle

Thermodynamics of a Combined Cycle [1/2]


In simple cycle application, thermal efficiency increases with the pressure ratio at a given firing

temperature.

For a given pressure ratio, thermal efficiency decrease as the TIT increases. This is because the lager

amount of cooling air for turbine blade is required as the TIT increases.

The pressure ratio resulting in maximum output and maximum efficiency change with firing temperature.

The higher the pressure ratio, the greater benefits from the increased firing temperature.

The power increases with the firing temperature at a given pressure ratio. However, efficiency decrease

because the flow of the cooling air extracted increase with the firing temperature.

Simple cycle

Combined cycle

In combined cycle applications, pressure ratio have a less pronounced effect on efficiency.

As pressure ratio increases, specific power decreases.

Thermal efficiency increases with firing temperature.

The optimum cycle parameters for combined cycle are different from simple cycle.

Simple cycle efficiency is achieved with high pressure ratios. However, combined cycle efficiency is

obtained with more modest pressure ratios and greater firing temperatures.

Thermodynamics of a Combined Cycle [2/2]


Efficiency of Combined Cycle [1/3]


SFGT

STGTCC

qq

ww

CC: gross efficiency of the combined cycle

wGT: output of gas turbine

wST: output of steam turbine

qGT: heat input in the gas turbine

qSF: heat input through supplementary firing in the HRSG

SFGT

AuxSTGTnetCC

qq

www

,

CC,net: net efficiency of the combined cycle

wAux: auxiliary consumption (=station service power consumption + electrical

losses)



The efficiency of the simple cycle gas turbine,

GT

GTGT

q

w

The efficiency of the simple cycle steam turbine,

SFExhGT

STST

qq

w

,

GTGTExhGT qq 1,

SFGTGT

STST

qq

w

1

The efficiency of the combined cycle without supplementary firing,

GT

GTGTSTGTGTCC

q

qq

1

GTSTGTCC 1



[Exercise 4.2]

Calculate the combined cycle efficiency. The efficiencies are 26 and 33% for a gas turbine cycle and a steam

turbine cycle, respectively.

[Solution]

The efficiency of the combined cycle is

[Note]

This calculation is based on the assumption of no pressure drops at the various cycle locations. Therefore,

this cycle efficiency represents an optimistic estimate. When the pressure drops are taken into account, the

efficiency of the combined cycle is greatly reduced. When detailed design of a combined cycle plant is made,

it usually shows the plant efficiency in the range of 38 to 42%.

GTSTGTCC 1

504.0CC

Efficiency of Combined Cycle


When the operation flexibility is important, such as marine applications, a mechanically independent power

turbine is used.

Compressor and high pressure turbine combination acts as a gas generator for the low pressure turbine.

Fuel flow to the combustor is controlled to achieve variation of power. This will cause a decrease in cycle

pressure ratio and maximum temperature.

At off-design conditions the power output reduces with the result that the thermal efficiency deteriorates

considerably at part loads.

Compressor

Fuel Combustor

LP Turbine

Air

Power

Exhaust gas

HP Turbine

Two-Shaft GT


Closed Cycle


Closed Cycle

In a closed cycle, the working fluid is continuously recirculated. It may be air or another gas such as helium.

Usually, the gas turbine is of intercooled recuperated configuration.

However, the combustor is replaced by a heat exchanger as fuel can not be burnt directly.

The heat source for the cycle may be a separate combustor burning normally unsuitable fuels, such as coal,

a nuclear energy, etc.

On leaving the recuperator, the working fluid must pass through a pre-cooler where heat is rejected to an

external medium, such as sea water, to return it to the fixed inlet temperature, usually between 15C and

30C.

The pressure at inlet to the gas turbine is maintained against leakage from the system by an auxiliary

compressor supplying a large storage tank called an accumulator.

The high density of the working fluid at engine inlet enables very high power output, which is the main

benefit of the closed cycle.


Factors Affecting Gas Turbine Performance


Generals

The gas turbine is a standardized machine, and can be used under widely different ambient conditions.

Manufacturers quote gas turbine performances at ISO ambient conditions of 15C(59F), 1.013 bar (14.7

psia), and 60% relative humidity.

Gas turbine performance is mainly governed by pressure ratio, turbine inlet temperature, and efficiency of

each parts.

The performance of gas turbine is affected by its inlet and exit conditions. The most important items are

pressure and temperature. Ambient weather conditions are the most obvious changes.

Since the gas turbine is an air-breathing machine, its performance is changed by anything that affects the

density and/or mass flow rate of air intake to the compressor.

A smaller weight of air requires a smaller weight of fuel to mix with, and the mixture then produces less

power when burned.

Most peak power enhancement opportunities exists in the topping cycle.

In general, however, performance enhancements to the gas turbines will carry with them an increase in

bottoming cycle performance due to an associated increase in gas turbine exhaust energy.

Duct firing within HRSG is an exceptional performance enhancement occurred in the bottoming cycle.


Factors Affecting GT Performance

Factors to be considered individually

Is there a need for peak power production with

premium paid for the resulting power?

Does peak power demand occur on hot days

(summer peaking) only?

Is there a need to compensate the power

reduction continuously during summer period?

Is frequency support required?

Solutions for power augmentation

Supplementary firing in HRSG

Steam / water injection

GT peak load firing

GT inlet air cooling


Options for power enhancementsTypical performance impact

Output Heat Rate

Base configuration Base Base

Evaporative cooling GT inlet air (85% effective cooler) +5.2 % -

Chill GT inlet air to 45F +10.7 % +1.6 %

GT peak load operation +5.2 % 1.0 %

GT steam injection (5% of GT airflow) +3.4 % +4.2 %

GT water injection (2.9% of GT airflow) +5.9 % +4.8 %

HRSG supplementary firing +28 % +9 %

Note: 1. Site conditions = 90F, 30% RH(Relative Humidity)

2. Fuel = NG

3. 3-pressure, reheat steam cycle

4. At sites where large power enhancement is possible, the owner must verify that the

added power is within the capabilities of the generator and transformer

Output = m h•

Various Options for Power Enhancement


1. Ambient Temperature [1/7]

Pe

rce

nt

de

sig

n

130

120

110

100

90

80

700 20 40 60 80 100 120 F

-18 -7 4 16 27 38 49 C

Compressor inlet temperature

The output and thermal efficiency of

the gas turbines decrease as air

temperature increases. This is because

an air density decreases as the

ambient air temperature increases,

thus the mass flow rate of air

decreases because industrial gas

turbines running at constant speed are

constant volume flow machines.

The thermal efficiency decreases as

the air temperature increases. This is

because compressor driving power

increases as the air temperature

increase. In addition, heat transfer

efficiency of the blade cooling system

decreases as the air temperature

increases.



The specific power consumed by the compressor increases

proportional to the inlet air temperature (in K) without a

corresponding increase in the turbine output.

The exhaust gas temperature increases as the inlet air

temperature increases because the turbine pressure ratio is

reduced, although the gas turbine inlet temperature remains

constant. This is the main reason for that the gas turbine

output and efficiency decrease while the ambient air

temperature increases.

However, the effect on the performance of the combined cycle

is somewhat less because a higher exhaust gas temperature

improves the performance of the steam cycle.

T

s

1

2

3

4

1

2

3

4

1

1

1,

CPR

Tcw

C

op

C


Effect on Combined Cycle Efficiency


Rela

tive e

ffic

iency,

%

Gas turbine

100

0

Air temperature, C

105

95

10 20 30 40-10

Based on constant

condenser pressure

Steam turbine

Combined cycle

An increase in the inlet air temperature has a

slightly positive effect on the efficiency of the

combined cycle plant, while other ambient

conditions as well as condenser pressure

remain constant.

Because the increased gas turbine exhaust gas

temperature improves the efficiency of the

steam process, it more than compensates for

the reduced efficiency of the gas turbine unit.

According to the open literature, with each one-

degree temperature increase above 30°C, power

output of the gas turbines drops by 0.50%–

1.02% while efficiency drops by approximately

0.24%. Steam turbine power output and

efficiency are not significantly changed by

changing air temperature, while net CCGT power

output drops by 0.3%–0.6% and net efficiency

drops by approximately 0.01% per degree above

30°C.

Source: Kehlhofer et al., 2009


Effect on Combined Cycle Efficiency


Com

bin

ed c

ycle

effic

iency,

%

Ambient air temperature, K


Net efficiency of a combined cycle power plant as a function of river water temperature.



Effect on Combined Cycle Output

The power output of the combined cycle

decreases as the inlet air temperature

increases.

In a combined cycle plant, gas turbines

contribute approximately two-thirds of the

power production, while the steam turbine

contributes the remaining one-third.

The combined cycle power output curve is

dominated by the gas turbine output curve, and

it is expected that changes in air temperature

will have more significant impact on plant

power output than changes in water

temperature.

The power output of the combined cycle is

affected differently from the efficiency because

change in mass flow of inlet air and exhaust

gases are more dominant than the exhaust gas

temperature.

Source: Kehlhofer et al., 2009


Rela

tive p

ow

er

outp

ut, %

Gas turbine

100

0

Air temperature, C

120

90

10 20 30 40-10

Based on constant

condenser pressure

110



When the ambient temperature is low, gas turbine output and HRSG steam production are increased above

plant rating point.

Condenser (exhaust) pressure directly influenced by ambient air or cooling water temperature.

Condenser pressure is expected to be lowest at low ambient air / cooling water temperature, and exhaust

annulus velocity will be the highest.


2. Ambient Air Pressure

Gas turbine performance is quoted at an air

pressure of 1.013 bar – ISO conditions, which

corresponds to the average pressure

prevailing at sea level.

A different site elevation and daily weather

variations result in a different pressure.

The air density reduces as the site elevation

increases. Therefore, airflow and output

decrease as the site elevation increases.

However, the air pressure has no effect on the

efficiency if the ambient temperature is

constant, even though the output decreases

as the pressure decreases. This is because

the backpressure of the gas turbine is

correspondingly lower at a lower ambient

pressure.

GT ModelCC Configu-

ration

Ambient

Temp.,CSite

Site

Elevation, m

CC Thermal

Effcy., %

CC Net

Power, MW

GT Net

Power, MW

ST Net

Power, MW

PG7221FA 2-on-128.1

(82.5F)

Las Vegas 664 53.4 437.0 285.3 151.7

Miami Sea side 53.2 475.5 312.9 162.6

Altitude x 103 feet

20 4 6 80.5

0.6

0.7

0.8

0.9

1.0

Correction

factor

Atmospheric

pressure

11.0

12.0

13.0

14.0

15.0

Co

rrection f

acto

r

Atm

osph

eric p

ressure

, p

sia


3. Humidity

Humid air is less dense than dry air.

In the past, this effect was thought to be too small to

be considered.

However, as the size of gas turbine increases, this

effect become important.

Steam or water injection for NOx control makes this

effect more significance.

Specific humidity (kg water vapor/kg dry air)

ISO specific humidity

0.0064

60% RH

1.008

0.996

Co

rre

ctio

n f

acto

r

0.010 0.02 0.03

1.006

1.004

1.002

1

0.998

0.994

1.010


4. Inlet & Exhaust Pressure Drop [1/5]

Inlet filter

Evaporative cooler or chiller

Anti-icing system

Silencer (The large frontal areas of the

compressors reduce the inlet velocities, thus

reducing air noise)


Hot-end drive Cold-end drive

• In the hot-end drive configuration, the output shaft

extends out the rear of the turbine.

• The designer is faced with many constraints, such

as output shaft length, high EGT, exhaust duct

turbulence, pressure drop, and maintenance

accessibility.

• Insufficient attention to any of these details, in the

design process, often results in power loss,

vibration, shaft or coupling failures, and increased

down-time for maintenance.

• This configuration is difficult to service as the

assembly must be fitted through the exhaust duct.

• In the cold-end drive configuration, the output shaft

extends out the front of the compressor.

• The single disadvantage is that the compressor

inlet must be configured to accommodate output

shaft.

• The inlet duct must be turbulent free and provide

uniform, vortex free, flow over the all operating

range.

• Inlet turbulence may induce surge in the

compressor resulting in complete destruction of the

unit.

MS7001E, GE MS7001F, GE



Inlet Pressure Drop

The improved operating performance associated with a

lower inlet air velocity design must be evaluated against the

associated higher capital cost.

A similar cost evaluation determines the optimum point that

dirty air filters, which have higher pressure losses, should

be changed out.

Inlet pressure drop is a function of the inlet air system design and cleanliness of the inlet air filters.

Lower inlet air pressure losses can be achieved by designing for lower inlet air velocities through the filter,

silencer, and duct.


Inlet pressure prop, inH2O

Co

rre

ctio

n f

acto

r

0 1 2 3 4 5 6 7 8 9 100.96

0.97

0.98

0.99

1.00

1.01

1.02


Exhaust Pressure Drop

Higher exhaust pressure loss is primarily a function of

the exhaust system design.

For a simple cycle applications, the exhaust system

typically consists of an exhaust duct, silencers, and a

stack.

Exhaust pressure losses of 4.0 to 5.0 inH2O are typical

for simple cycle gas turbines.

For combined cycle or cogeneration applications, the

exhaust gases pass through an HRSG with the

associated additional.

Exhaust pressure losses of 10 to 17 inH2O are typical

for combined cycle and cogeneration applications

depending on the complexity of the cycle arrangement,

exhaust emission control, or noise-abatement.


Exhaust pressure drop, inH2O

Co

rre

ctio

n f

acto

r1

1.020

1.015

1.010

1.005

0.995

0.990

0.985

0.980

2 3 4 5 6 7 8 9 10 11 12 13 140 15

1.025

1.000

0.975


MS7001EA 기준

4.0 inH2O (10mbar) inlet pressure drop produces:

1.42% Power output loss

0.45% Heat rate increase

1.1C Exhaust temperature increase

4.0 inH2O (10mbar) Exhaust pressure drop produces:

0.42% Power output loss

0.42% Heat rate increase

1.1C Exhaust temperature increase

Inserting air filter, silencer, evaporative coolers or chillers into the inlet or heat recovery devices in the

exhaust causes pressure losses in the system.

The effects of these pressure losses are unique to each gas turbine models. This is because the amount of

pressure drop at the exit of compressor is pressure drop at the inlet times pressure ratio.

Hot-end drive has not been used since the cold-end drive type gas turbines have developed.

HRSG flue gas draft losses: approximately 25 mbar, 35 mbar if catalysts are required.


Hot-end drive (“E” technology)


5. Fuel [1/7]

Fuel affects combined cycle performance in a variety of ways.

Output of the gas turbine can be defined as the product of mass flow, specific heat, and temperature

differential across the turbine. Here, specific heat (cp) means that the heat energy in the combustion

products.

The mass flow in this equation is the sum of compressor air flow and fuel flow.

Natural gas (methane) produces nearly 2% higher output than does distillate oil. This is because of the

higher specific heat in the combustion products of natural gas, resulting from the higher water vapor content

produced by the higher hydrogen/carbon ratio of methane. This effect is noted even though the mass flow of

natural gas is lower than that of distillate oil. Here the effects of specific heat were greater than and in

opposition to the effects of mass flow rate.

4343 TTcmhhmW pT

Model FuelISO base rati

ng, kW

Heat rate,

Btu/kWh

Exhaust flow,

kg/hr x10-3

EGT,

℃Pressure

ratio

PG7251FB

N.G. 184,400 9,245 1613 623 18.4

D.O. 177,700 9,975 1677 569 18.7


C + O2 = CO2 + 33.9 MJ/kg

H2 + 1/2O2 = H2O(water) + 143.0 MJ/kg (HHV)

H2 + 1/2O2 = H2O(vapor) + 120.6 MJ/kg (LHV)

S + O2 = SO2 + 9.28 MJ/kg

5. Fuel [2/7]


The Composition of Natural Gases

The composition on a molar basis of natural gases is as follows:

The average heat content of natural gas is 1,030 Btu/ft3 on an HHV basis and 930 Btu/ft3 on

an LHV basis – about a 10% difference.

Composition, mol% A B C D E F

Methane

Ethane

Propane

Isobutane

Normal butane

Isopentane

Normal pentane

Hexane

Nitrogen

Carbon dioxide

Hydrogen sulphide

Heating value, Btu/ft3

95.0

1.9

0.5

0.5

0.1

0.1

0.1

0.1

1.5

0.2

0.0

?

94.3

2.1

0.4

0.0

0.2

0.0

0.0

0.0

0.0

0.0

2.8

1010

72.3

5.9

2.7

0.2

0.3

0.0

0.2

0.0

0.0

17.8

0.1

934

88.9

6.3

1.8

0.1

0.2

0.0

0.0

0.0

0.0

2.2

0.1

1071

75.4

6.4

3.6

0.6

1.0

0.2

0.1

0.0

0.0

12.0

0.1

1044

85.6

7.8

1.4

0.1

0.0

0.1

0.0

0.0

0.0

4.7

0.2

1051

5. Fuel [3/7]


Plant output and efficiency can be reduced when the fuels containing higher sulfur content are used. This is

because higher stack gas temperature is required to prevent condensation of corrosive sulfuric acid.

Plant output and efficiency can be reduced when the ash bearing fuels (crude oil, residual oil, blends, or

heavy distillate) are used because of fouling occurred in gas turbine and HRSG.

Heavy fuels normally cannot be ignited for gas turbine startup; therefore a startup and shutdown fuel, usually

light distillate, is needed with its own storage, forwarding system, and fuel changeover equipment.

The LHV of the fuel is important because it defines the mass flow of fuel supplied to the gas turbine.

The lower the LHV, the higher the mass flow of fuel required to provide a certain chemical heat input,

normally resulting in a higher power output and efficiency. However, there is no clear relationship between

fuel lower heating value and output.

This is why low BTU gases can result in high power outputs if they are supplied at the pressure required by

the gas turbine.

This effect is noted even though the mass flow of methane is lower than the mass flow of distillate fuel.

Here the effects of specific heat were greater than that of mass flow.

5. Fuel [4/7]


Ash deposition

on turbine

vanes

5. Fuel [5/7]

Degradation in CCPP after 8,000 hours of operation

Clean fuel Heavy or crude oil

Plant output, % 0.8~1.5 4.0~5.5

Plant efficiency, % 0.5~0.8 1.5~1.9


Hot corrosion of blades Burned turbine blades

In the past, corrosion is one of the major causes of gas turbine failures.

Corrosion problems have been eliminated by the use of advanced materials and coating.

Whenever heavy fuels are used, particularly those containing vanadium or sodium, it is necessary to use

additives or treat the fuel to prevent high-temperature corrosion.

The additives commonly used are based on magnesium, chromium, or silicon.

5. Fuel [6/7]


Effects of Fuel Heat Value on Output

As the amount of inert gas is increased, the decrease in

LHV will provide an increase in output.

This is the major impact of IGCC type fuels that have

large amounts of inert gas in the fuel.

This mass flow addition, which is not compressed by the

gas turbine’s compressor, increase the turbine output.

5. Fuel [7/7]


6. Fuel Heating [1/2]

One way of improving the cycle efficiency is to

raise the apparent LHV (LHV + sensible heat) of

the fuel by preheating it with hot water from the

IP economizer of the HRSG.

Heated fuel gas gives higher turbine efficiency

because of the reduced fuel flow required to

raise the total gas temperature to firing

temperature.

Fuel heating will result in slightly lower gas

turbine output (almost negligible) because of the

incremental volume flow decrease.

The reduction in combined cycle output is

typically greater than simple cycle output

because energy that would otherwise be used to

make steam.

Actual combined cycle output and efficiency

changes are dependent on fuel temperature rise

and cycle design.

G

Fuel gas

ST

Condenser

G

Air

Stack gas

HRSG


For combined cycle applications, fuel temperatures on the order of 150 to 230°C (300~450°F) are generally

economically optimal.

Provided the fuel constituents are acceptable, fuel temperatures can potentially be increased up to

approximately 370°C(700°F) before carbon deposits begin to form on heat transfer surfaces.

Typical F-class three-pressure reheat systems use water from the intermediate pressure economizer to

heat the fuel to approximately 185°C (365°F). Under this conditions, efficiency gains of approximately 0.3

points can be expected for units with no stack temperature limitations.

Another factor is the gas supply pressure, depended on the combustor design and the gas turbine pressure

ratio.

If the gas turbine pressure ratio is high, a gas compressor may be required to increase fuel pressure. In this

case, the temperature of the fuel is increased in proportion to the pressure ratio and the benefit of gas

preheater will be reduced.

It is important to ensure that the fuel does not enter the steam system because maximum steam

temperatures are typically above the auto ignition temperature for gas fuels.

For a system utilizing a direct water-to-fuel heat exchangers, the water pressure is maintained above the

fuel pressure so that any leakage takes place in the fuel system.

Additional system design and operation requirements ensure that the fuel does not enter the steam system

during periods when the water system is not pressurized.

6. Fuel Heating [2/2]


7. Steam Injection [1/6]

Effects of Steam Injection on Output and Heat Rate [MS7001EA]

Compressor Inlet Temperature

Options for Power

Enhancements

Performance Impact

Output Heat

Rate

Base configuration Base Base

Evaporative cooling GT inlet air

(85% effective cooler)+5.2 % -

Chill GT inlet air to 45F +10.7 % +1.6 %

GT peak load operation +5.2 % 1.0 %

GT steam injection (5% of GT

airflow)+3.4 % +4.2 %

GT water injection (2.9% of GT

airflow)+5.9 % +4.8 %

HRSG supplementary firing +28 % +9 %

Note: 1. Site conditions = 90F, 30% RH(Relative Humidity)

2. Fuel = NG

3. 3-pressure, reheat steam cycle



Re

lative

po

we

r o

utp

ut, %

Water or steam/fuel ratio,

114

Rela

tive

effic

iency,

%112

110

108

106

104

102

100

98

96

940.5 1.0 1.5

Hot water injection (150C)

0.0


Diluent injection is accomplished by admitting water or steam in the cap area or head-end of the

combustion liner to reduce the peak flame temperature.

Actually, this has been used for NOx control to meet environmental regulation.

The mass flow passing through the gas turbine increase with the amount of water or steam injection.

Increased mass flow produces higher power output.

Generally, the amount of water is limited to the amount required to meet the NOx abatement in order to

minimize operating cost and impact on inspection intervals.

When steam is injected for power augmentation, it can be introduced into the compressor discharge casing

of the gas turbine as well as combustor.

Normally, gas turbines are designed to allow up to 5% of the compressor airflow for steam injection.

Steam must contain 50F(28C) superheat and be at pressures comparable to fuel gas pressures (at least

40 bar above the compressor discharge).

The way steam is injected must be done very carefully so as to avoid compressor surge.

Gas turbine output and heat rate increase 3.4% and 4.2% respectively, by the steam injection of 5% of the

compressor airflow.

Water or steam injection for emission control or power augmentation can impact parts lives and

maintenance intervals.



The control system on most base load

applications reduces firing temperature as

water or steam is injected. This is known as

dry control curve operation.

The dry control curve operation counters

the effect of higher heat transfer on the gas

side, and results in no net impact on bucket

life.

This is the standard configuration for all gas

turbines, both with and without water or

steam injection.

Dry Control

On some installations, however, the control system is

designed to maintain firing temperature constant with

water or steam injection level. This is known as wet

control curve operation.

The wet control curve operation results in additional unit

output, but decreases parts life.

Units controlled in this way are generally in peaking

applications where annual operating hours are low or

where operators have determined that reduced parts

lives are justified by the power advantage.

Wet Control

An additional factor associated with water or steam injection relates to the higher aerodynamic loading on

the turbine components that results for the injected water increasing cycle pressure ratio.

This additional loading can increase the downstream deflection rate of the second- and third-stage nozzles,

which would reduce repair interval for those components.



EGT Control Curve – MS7001EA

[ Steam Injection for 25 ppm NOx ]

GER-3620K


Wet control 3% steam injection

TF = 2020F(1104C)

Load ratio = 1.10

3% steam injection

TF = 1994F(1090C)

Load ratio = 1.08

0% steam injection

TF = 2020F(1104C)

Load ratio = 1.0

EG

T, F

Compressor discharge pressure, psig

Dry control

The wet control maintains

constant TF


Steam/water injection increases metal temperature of hot-gas-path components in the case

of constant firing temperature operation.

• Water affects gas transport properties:

k – thermal conductivity

cp – specific heat

– viscosity

• This increases heat transfer coefficient, which increases metal temperature and

decreases bucket life

Example (MS7001EA 1st stage bucket):

• 3% steam injection (25 ppm NOx)

• h = +4% (heat transfer coefficient)

• Tmetal = +15F (8C)

• Life = – 33%



8. Air Extraction

Effects of Air Extraction on Output and Heat Rate

In some gas turbine applications, it may be desirable to

extract air from the compressor.

In general, up to 5% of the compressor airflow can be

extracted from the compressor discharge casing without

modification to casings or on-base piping.

Air extraction between 6% and 20% may be possible,

depending on the machine and combustor, with some

modification to the casings, piping and controls.

Air extractions above 20% will require extensive

modification to the turbine casing and unit configuration.

As a rule of thumb, every 1% in air extraction results in a

2% loss in power. Ambient temperature%

Effect

on o

utp

ut

C

F1200 20 40 60 80 100

0 -7 4 16 27 38 49

100

80

60

40

20

0

-20

-40

-60

% E

ffect

on h

eat

rate

20%

15%

10%

5%

5%

10%

15%

20%


9. Inlet Air Cooling [1/24]

Roughly, 1C temperature decrease corresponds to a combined cycle power increase of about +0.4 to

0.5% and overall efficiency remains more or less same.

Compressor

Fuel gas

pre-heater

Turbine

Fuel oil

Fuel gas

compressor

Fuel oil

treatment

Fuel gas

Water

Evaporative

cooler

or

chiller

Wet compression

(Overspray)

Compressor

washing

Fuel additives

Water

(NOx reduction,

power augmentation)

Cooling air

cooler

Air

Air filter

Drain

Water

Steam

Water

Inlet Fogger

(spray cooler)


건구온도 vs. 습구온도

상대습도(relative humidity): 공기중에 있는 수증기의 양과 그때의 온도에서 공기중에 최대로 포함할 수 있는 수증기의 양을백분율로 표현한 값.

건구온도(dry bulb temperature): 일반 온도계 측정한 온도.

습구온도(wet bulb temperature): 온도계 아래 부분 동그란 구면을 거즈로 감싸고 거즈의 한쪽 끝을 물이 담긴 그릇에 넣어그릇에서 빨아올린 물이 끊임없이 온도계의 구면에서 증발하도록 한 상태에서 측정한 온도. 물이 증발하면서 기화열을 빼앗아 가기 때문에 건구온도보다 더 낮은 온도를 나타냄.

공기중의 습도가 낮으면 물이 더 많이 증발할 수 있어서 열을더 많이 빼앗아 가기 때문에 건구와 습구온도 차이가 더 커짐.

일반적으로 건구온도와 습구온도 차이에 의해서 습도를 계산.

습도 계산표가 있어서 건구온도와 습구온도를 알면 그 때의 습도를 찾을 수 있음. 대부분 습도 계산표를 이용하여 습도 확인.

[ 습구온도계 ]



For applications where significant power demand and highest electricity prices occur during the hot summer,

a gas turbine air inlet cooling system is a useful option for increasing power output.

Inlet air cooling increases output because the mass flow rate of air passing through the compressor

increases as air temperature decreases.

A decrease in the inlet dry-bulb temperature by 10F(5.6C) will normally result in around 2.7% power

increase of a combined cycle using heavy-duty gas turbines.

The output of the simple-cycle gas turbines is also increased by the same amount.

There are three basic systems currently available for inlet air cooling.

• The first and perhaps the most widely used system is evaporative cooler. Evaporative coolers use the

high efficiency evaporative media for the evaporation of water to decrease the gas turbine inlet air

temperature.

• The second one is a fogger system, also called as spray cooler. This is classified as evaporative cooling.

• The third system employs various ways to chill the inlet air. In this system, the coolant (usually chilled

water) flows through a heat exchanger located in the inlet duct to remove heat from the inlet air.

Evaporative cooling is limited by the wet-bulb temperature.

Chilling, however, can cool the inlet air to temperature that are lower than the wet-bulb temperature, thus

providing additional output, although chilling is much more expensive.

Depending on the combustion and control system, evaporative cooling may reduce NOx emissions; however,

this is very little because of current dry low NOx technology.



High efficiency evaporative media

Evaporative Cooler (Wetted Honeycomb Evaporative Coolers)




Conventional evaporative coolers use a wetted honeycomb type medium to maximize evaporative surface

area and the cooling effectiveness.

The medium for gas turbines is typically 12 inches thick and covers the entire cross-section of the filter

house or the inlet air duct.

The pressure drop caused by evaporative media and droplet eliminator is 1 inH2O.



The plant output and efficiency decrease due to this pressure drop.

The reduction in gas turbine and combined cycle output is 0.35% and 0.3%, respectively.

A controller is provided to prevent operation of the evaporative cooler system below 60F(15.6C).

Icing could form if the system is allowed to operate below this temperature.

The whole system must be deactivated and drained to avoid damage to the water tank and piping if the

ambient temperature is expected to fall below freezing.

Evaporative cooling is a cost-effective method to recover capacity during periods of high temperature and

low or moderate relative humidity.

Evaporative cooling works on the principle of reducing the temperature of an air stream through water

evaporation.

The process of converting the water into a vapor state requires energy.

This energy is drawn from the air stream. The result is cooler, denser air.

There are limitations that must be considered for each site condition. The key design parameters are the wet

and dry bulb temperature and the allowable load limits for the generator and the transformer.

At sites where large reductions in the compressor inlet temperature are possible, the owner must verify that

the added power is within the capabilities of the generator and transformer.




Evaporative Cooling - Theory

Theoretically, the lowest temperature that can be achieved by adding water to the air is equal to the ambient

wet-bulb temperature.

Practically, however, this level of cooling is difficult to achieve.

The actual temperature drop realized is a function of both the equipment design and atmospheric conditions.

Other factors being constant, the effectiveness of an evaporative cooling system depends on the surface

area of water exposed to the air stream and the residence time.

T means air temperature. Subscripts 1 and 2 refer to inlet and exit of the cooler, respectively. Subscripts DB

and WB refer to ‘dry bulb’ and ‘wet bulb’, respectively.

Temperature drop of the compressor inlet air is proportional to the difference between wet and dry bulb

temperature. If the effectiveness is 85%, the temperature drop is

The effectiveness of evaporative cooler is typically 85% and of foggers somewhat higher at 90 to 95%.

WBDB

DBDB

TT

TTessEffectivenCooler

,2,1

,2,1

WBDB TTdropeTemperatur ,2,185.0



Psychrometric Chart [습도 선도]

Water

Evaporated

Degrees Cooled

Dry Bulb Temperature

[Solution]

The corresponding wet-bulb

temperature is 70F.

T = 0.85(100-70) = 25F (14C)

[Example 9.1]

Ambient temperature is 100F (37.8C)

and relative humidity is 20%. Calculate

the temperature drop through the cooler.

The effectiveness of the evaporation

system is 85%.



Evaporative Cooling

Roughly, 1C temperature decrease corresponds to a

combined cycle power increase of about 0.4 to 0.5% and

overall efficiency remains more or less same.

The exact increase in power available from a particular gas

turbine as a result of evaporative cooling depends on the

machine model and site altitude, as well as on the ambient

temperature and humidity.

However, the chart given in the figure can be used to get the

power increase from evaporative cooling.

As would be anticipated, power increase is greatest in hot, dry

weather.

Evaporative cooling is limited to ambient temperatures (15C)

and above (compressor inlet temperature >7.2C) because of

the potential for icing the compressor.

An evaporative cooling does only make sense at locations

with humidity below 70 to 80%.



Gas turbines have been used foggers, also called as spray coolers, since mid-1980s.

These systems atomize the supply of water into billions of tiny droplets.

The droplets require a certain amount of residence time in the air stream to evaporate. The size of droplet

plays an important role in determining the surface area of water exposed to the airstream and, therefore, to

the speed of evaporation.

The water droplets should be atomized to less than 20 m in foggers.

Fogger

Demineralized water is used

to reduce compressor fouling

or nozzle plugging. However,

it necessitates the use of a

high grade stainless steel for

all wetted parts.



Fogger - EPRI Spray Nozzle Array

A typical spray-impingement

fog nozzle

Nozzle fog spray pattern



Two methods are used for water atomization.

The first relies on compressor air in the nozzles to atomize the water.

The second uses a high pressure pump to force the water through a small orifice.

Air-atomized nozzles require less water pressure.

However, they result in low power output due to the air extraction from the gas turbine. An air-atomized

system using compressor discharge air would reduce the power output 1.3% (EPRI, TR-104612).

The typical air-to-water mass ratio is 0.6 (volume ratio is 500).

Some high-pressure pumps use swirlers to break the water into small droplets.

Other force the water on an impingement pin to generate the same effect.

A typical high-pressure pumped fog system has an operating pressure of between 1000 and 3000 psi (6.8

and 20.4 MPa).

Fogger



Careful application of these systems is

essential, because condensation or

carryover of water can be causes of severe

compressor fouling and performance

degradation.

These systems generally are followed by

moisture separators or coalescing pads to

reduce the possibility of moisture carryover.

More spray flow was removed (~70%) by the

Large Droplet Eliminator than was originally

anticipated (~58%) in the EPRI test (TR-

108057).

Normally, water droplets are agglomerated

by turbulent fluctuations and become large

droplets.

Analysis of the drain water gives some

beneficial air scrubbing effects when the

spray cooler is operating.

Fogger - Large Droplet Eliminator

A large droplet eliminator (LDE) is installed in the in the inlet

housing downstream of the spray nozzle array to remove

large water droplets from the air stream. The modules are

manufactured using polypropylene with sine curve shaped

vanes.



The power increase from evaporative cooling is about 3.5% for every 10F (5.6C) of cooling.

Evaporative cooling is limited by the difference between the dry bulb and wet bulb temperatures.

If sufficient water can be introduced into the air such that the air becomes fully saturated, the air

temperature will be reduced to the wet bulb temperature.

The amount of cooling is limited by the potential for icing as the air flow speeds up in the bellmouth and the

static air temperature drops.

The icing limit is engine dependent but typically varies from 40F to 50F.

Fogger - Summary of ERPI Test Results



Evaporative Cooling

Evaporative Cooler Fogger (Spray cooler)

Advantages

• Water quality requirements are less

severe than fogger system.

• Simple and reliable.

• More operating experience.

• Gas turbine inlet pressure drop is lower

than that of evaporative cooler and

provides increased output.

• Higher effectiveness.

• Potential for lower uprate costs and faster

installation time due to reduced duct

modifications compared to evaporative

cooler.

Disadvantages

• Uprates frequently require substantial

duct modifications.

• Higher gas turbine inlet pressure drop

than fogger system degrades output and

efficiency when not in use.

• Lower cooling effectiveness.

• Requires demineralized water.

• Higher parasitic load than evaporative

cooler for high-pressure pumped systems.

• Lower power increase for air-atomized

systems.

• Controls are more complex.



Compressor Turbine

Overspray (Wet compression)

Water

Evaporative

cooler

or

chiller

Overspray

(wet compression)

Compressor

washing

Cooling air

cooler

Air

Air filter

Drain

Water

Steam

Water

Inlet Fogger

(spray cooler)

• Overspray is defined as the excess spray beyond that

which is required to completely saturate the air.

• As an extension of the fogger system, water droplets

are allowed to enter the compressor and evaporation

takes place within the compressor.

• Droplets are evaporated inside the compressor to give

evaporative intercooling effect.

Overspray



LM6000 Sprint

Overspray

GER-3620K

Overspray (Wet compression)

• The power increase resulting from overspray is about

5% for every 1% overspray (overspray water mass is

expressed as a percentage of inlet air mass).

• The amount of overspray will depend on ambient

conditions.



Overspray

GER-3620K



There are two types of inlet chilling systems, direct chillers and thermal storage.

Liquefied natural gas (LNG) systems use the cooling generated by the vaporization of liquefied gas in the

fuel supply.

Thermal storage systems use off-peak power to store thermal energy in the form of ice.

During peak power periods, the ice is used to perform inlet chilling.

Direct chilling systems use mechanical or absorption chillers.

All these options can be installed in new plants or retrofitted in older plants.

The chilling achieved by using cooling coils depends on the design of the equipment and ambient

conditions.

Unlike evaporative coolers, cooling coils are capable of lowering the temperature below the wet-bulb

temperature.

The capacity of the inlet chilling device, the compressor’s acceptable temperature and humidity limits, and

the effectiveness of the coils limit actual reduction in temperature.

Figure illustrates a typical cooling process from an ambient dry-bulb temperature of 100F(37.8C) and 20%

relative humidity.

The initial cooling process follows a line of constant specific humidity.

Chiller



As the air approaches saturation,

condensation starts to occur.

Additional cooling results in further

condensation.

Mist eliminator should be installed

downstream of coils to prevent

condensed water from entering the

gas turbine.

The air can be cooled below the

ambient wet-bulb temperature.

However, the compressor inlet

temperature should be higher than

45F(7.2C) with a relative humidity of

95%.

Icing will form at lower temperature,

resulting in possible equipment

damage.

Water

Evaporated

Degrees Cooled

Dry Bulb Temperature

Chiller



Chiller



Chiller

F-Class gas turbine inlet filter house, showing

installation of chiller coils. Coil manifolds are the

vertical pipes along side the filter house. This filter

house is “passively balanced” with a third “reverse

return manifold”.

Filter houses for chilling applications are much

larger than standard models. A larger “face area”

keeps pressure drop across the coils low. This filter

house also has a symmetrical transition duct that

improves the airflow across the coils.

The fogger systems react to the ambient weather conditions, being limited to the spread between DB and

WB. However, chiller systems break through the WB and dew-point barriers that would limit fogger systems.

The power output enhancement associated with chiller systems can be nearly twice that of the fogger

systems. In addition, temperature of the inlet air can be as constant as possible using chiller systems.



CaseInlet Air

Cooling

GT Output,

kW (each)

ST Output,

kW (gross)

Auxiliary

Power, kW

Net Plant

Output,

kW

Each GT

Fuel Input,

MMBtu/hr

(LHV)

Total Duct

Burner Fuel

Input,

MMBtu/hr

(LHV)

Net Plant

Heat Rate,

Btu/kWh

(LHV)

1 None 14207 169942 12173 452183 1440.373 0.000 6371

2 Fogger 159850 174642 12489 481853 1531.400 0.000 6356

3 Chiller 173833 178055 19799 505922 1632.324 0.000 6453

4 None 147185 252408 16575 530203 1440.300 654.636 6668

5 Fogger 159831 254283 16839 557106 1531.337 635.466 6638

6 Chiller 173813 255761 24071 579316 1632.256 619.586 6705

7 Chiller 164800 175366 15918 489048 1565.722 0.000 6403

• Model: STAG207FA

• Simulation software: GTPro & GTMaster

• Ambient conditions: 95F (35C), 40% RH

• Effectiveness of fogging system: 95%

• The chiller cools the inlet air temperature down to 50F (10C)

Summary for Performance Simulation Results



CaseInlet Air

Cooling

Duct

Firing

Net Plant

Output, kW

Incremental Output,

kW

Reference

Cost, M$

Incremental

Cost (to Case 1),

M$

Unit Cost,

S/kW

1 None No 452183 212.4 470

2 Fogger No 481853 29670 6.6% 215.1 2.7 446

3 Chiller No 505922 53739 11.9% 225.3 12.9 445

4 None Yes 530203 78020 17.3% 229.5 17.1 433

5 Fogger Yes 557106 104923 23.2% 231.5 19.0 416

6 Chiller Yes 579316 127133 28.1% 239.9 27.5 414

• All cost figures are provided by Thermoflow’s PEACE costing module.

• This software uses the plant configuration as provided by GTPro.

Summary for Capital Cost Simulation Results



Some gas turbine models can be operated at a higher firing temperature than their base rating.

This is called peak firing. During the peak firing operation, both simple-cycle and combined-cycle output will

increase.

Peak firing is available to get 3~10% higher output than the output at base load.

Normally, thermal efficiency of the plant is increased during peak firing of gas turbine because of higher

firing temperatures.

This mode of operation results in a shorter inspection interval and increased maintenance.

Despite this penalty, operating at elevated peak firing temperatures for short periods is cost-effective way

for power gain without any additional peripheral equipment.

“Peaking” at 110% rating will increase maintenance costs by a factor of 3 relative to base-load operation at

rated capacity, for any given period.

For an MS7001EA turbine, each hour of operation at peak load firing temperature (+100F/56C) is the

same, from a bucket parts life standpoint, as six hours of operation at base load.

10. Peak Firing [1/2]


Maintenance Factor

E Class

F ClassE Class

Peak Rating

Life Factor 6x

0 50 100 150

Firing Temperature, F

Ma

inte

na

nce

Fa

cto

r

100

10

6

0

10. Peak Firing [2/2]


11. Part Load Operation

Load, %30 40 50 60 70 80 90 100

65

60

75

70

85

80

95

90

100


12. Supplementary Firing in HRSG

It can be used to increase steam turbine capacity by

as much as 100%.

This will increase plant capacity by about 25%.

Cogeneration of power and process steam is usually

the incentive for HRSG supplementary firing.

There is a small performance penalty when operating unfired compared to operating a unit designed without

supplementary firing, and the magnitude of this performance penalty is directly proportional to the amount of

supplementary firing built into the combined-cycle plant.

The performance penalty is due to two factors: unfired operation results in lower steam flows and pressures

and, thus, lower steam turbine efficiency; also, the pumps, auxiliary equipment and generator are sized for

higher loads.

Operating unfired results in comparatively higher parasitic loads compared to a unit designed solely for

unfired operation.

Normally, thermal

efficiency of the plant is

decreased during HRSG

supplementary firing.


13. Cooling Water Temperature

T

s

2

P4’

1

1

2

3

a b

4

P4

a

4

The end pressure of steam expansion in the turbine is determined by the steam saturation

temperature depending on the cooling water temperature and heat transfer conditions on the

condenser tubes.


질의 및 응답

작성자: 이 병 은 (공학박사)작성일: 2015.02.11 (Ver.5)연락처: [email protected]

Mobile: 010-3122-2262저서: 실무 발전설비 열역학/증기터빈 열유체기술

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