Performance and Emission Characteristics of Natural Gas Combined Cycle Power Generation

System with Steam Injection and Oxyfuel Combustion

By

Nitin N. Varia

A Thesis Submitted in Partial Fulfillment of the Requirements for

the Degree of

Master of Applied Science in Mechanical Engineering

The Faculty of Engineering and Applied Science Program

University of Ontario Institute of Technology, Oshawa, Ontario, Canada

© Nitin N. Varia, 2014

ii

Certificate of Approval

iii

Abstract

Natural gas combined cycle power generation systems are gaining popularity due

to their high power generation efficiency and reduced emission. In the present work,

combined cycle power generation configuration systems studied with natural gas as

primary fuel oxidizes with air and pure oxygen separately. Steam is injected in main

combustion chamber and reheater combustion chamber individually and

simultaneously to understand the performance of combined cycle work output and

greenhouse gas emission. The effect of pressure ratio, turbine inlet temperature,

isentropic efficiency, ambient temperature on combined cycle work output, thermal

and exergy efficiency are carried out with and without steam injection. In present

range of investigation, it is observed that the steam injection increases gas cycle

efficiency and decreases the steam cycle efficiency. Ideal pressure ratio found to be

25 in all different combined cycle power generation system configurations. Maximum

CO2 emission reduction (7.2%) occurs when steam is injected in reheater

combustion chamber for fuel combustion with air and (3.2%) when steam injection

in both combustion chambers for oxy fuel configuration. Thermal efficiency of

combined cycle system increased by 8.2% when 10% steam injection in both

combustion chambers. In oxyfuel combustion, higher ratio of recycle flue gas brings

higher thermal efficiency and highest thermal efficiency achieved when steam is

injected in main combustion chamber only. Maximum exergy destruction found in

combustion chambers (57%), steam injection lowers exergy destruction by 4%.

More than 10% steam in combustion chamber brings combined cycle thermal

efficiency down.

Keywords: Steam Injection, Combined cycle, CO2 Capture, Oxyfuel, Turbines,

Energy efficiency, Exergy efficiency

iv

Acknowledgements

First and foremost, I am deeply grateful to my supervisor, Dr Bale V. Reddy for his

contributions of time and ideas to make my research experience productive and

stimulating. I sincerely appreciate his constant guidance and support during the

period of this work and most of all, his patience and understanding. It has been an

honor to work with them. I would also like to thank Dr. T. Srinivas for his time and

valuable suggestions. I would also take this opportunity to express my great

appreciations to the Faculty of Engineering and Applied Science at UOIT since

they have provided all the assistance and support to me.

Finally, I would like to thank my family, especially my parents and brother for all

their encouragement. And most of all for the loving support and patience of my

wife and sons throughout this research endeavor.

v

Table of contents

Certificate of Approval ............................................................................................ ii

Abstract .................................................................................................................. iii

Acknowledgements ................................................................................................ iv

Table of contents .................................................................................................... v

List of figures ....................................................................................................... viiii

List of tables .......................................................................................................... xii

Nomenclature ....................................................................................................... xii

Chapter 1: Introduction .......................................................................................... 1

1.1 Energy scenario ......................................................................................... 1

1.2 Natural gas ................................................................................................. 3

1.3 Carbon Dioxide Emission ........................................................................... 4

1.4 Combined cycle power plants .................................................................... 5

1.5 Gas Turbine with Steam Injection (STIG) .................................................. 8

1.6 Carbon capture mechanism and advances ............................................... 9

1.7 Oxy-fuel combustion ................................................................................ 11

1.8 Air separation unit (ASU) ......................................................................... 13

1.9 Exergy ...................................................................................................... 14

1.9.1 Exergy destruction ........................................................................... 15

1.10 Objective of the thesis work ................................................................... 15

Chapter 2: Literature review: Recent work .......................................................... 17

Chapter 3: Combine cycle power generation system configurations and

methodology ........................................................................................................ 26

3.1 Natural gas combined cycle configuration description............................. 26

3.1.1 Natural gas combined cycle power generation configuration 1 ...... 28

3.1.2 Natural gas combined cycle power generation configuration 2 ...... 29

3.1.3 Natural gas combined cycle power generation configuration 3 ...... 29

vi

3.1.4 Natural gas combined cycle power generation configuration 4 (oxy-fuel

cycle) ......................................................................................................... 30

3.1.5 Natural gas combined cycle power generation configuration 5 (oxy-fuel

cycle) ......................................................................................................... 30

3.1.6 Natural gas combined cycle power generation configuration 6 (oxy-fuel

cycle) ......................................................................................................... 31

3.2 Thermodynamic Analysis ......................................................................... 31

3.3 Methodology ............................................................................................. 32

3.3.1 Gas compressors (C1) & (C2) and intercooler (IC) ......................... 32

3.3.2 Primary and reheater combustion chambers CC1 and CC2 .......... 35

3.3.3 Natural gas fired gas turbines (GT1 and GT2) ................................ 44

3.3.4 Heat recovery steam generator (HRSG) ......................................... 46

3.3.4 Steam turbine (ST) .......................................................................... 47

3.3.5 Condenser (CON) ............................................................................ 48

3.3.6 Water circulation pump for bottoming cycle .................................... 49

Chapter 4: Results and discussions .................................................................... 51

4.1 Component range of values used in studies ............................................ 51

4.2 Effect of pressure ratio on combined cycle performance with fraction of steam

injection .......................................................................................................... 52

4.3 Effect on CO2 emission with fraction of steam injection .......................... 56

4.4 Effect of ambient temperature on combined cycle performance with fraction

of steam injection ........................................................................................... 57

4.5 Performance analysis on fuel ratio on combined cycle with fraction of steam

injection .......................................................................................................... 59

4.6 Effect of TIT on performance of combined cycle with fraction of steam

injection .......................................................................................................... 63

4.7 Effect of Isentropic efficiencies on performance of combined cycle with

fraction of steam injection .............................................................................. 68

vii

4.8 Effect of flue gas recycle on turbine inlet temperatures for configuration 4, 5

and 6 (oxyfuel combustion) ............................................................................ 70

4.9 Effect of steam injection on TIT with fraction of flue gas recycle for

configuration 4 (Oxyfuel combustion) ............................................................ 70

4.10 Effect of pressure ratio on combined cycle work output, efficiencies & CO2

emission for configurations 4, 5 and 6 (Oxyfuel combustion) ........................ 73

4.11 Exergy destruction in combined cycle system ....................................... 77

4.12 Result validation ..................................................................................... 78

Chapter 5: Conclusions ....................................................................................... 81

5.1 Principal contributions .............................................................................. 81

5.2 Conclusions .............................................................................................. 82

5.3 Recommendations ................................................................................... 83

References ........................................................................................................... 84

Appendix: EES CODE ......................................................................................... 87

viii

List of figures

Figure 1.1 Power generation comprises the largest source of CO2 emissions in

2007 [1] ….….….….….….….….….….….….….….….….….….….….….….….……..2

Figure 1.2 Canada electricity generation by fuel types [1] ….….….….….….………3

Figure 1.3 Electricity generations in Ontario by various energy sources [2] ……..4

Figure 1.4 Combined cycle operation with Brayton and Rankine cycle [18]…….....6

Figure 1.5 Combined cycle power plant with reheater combustion chamber

[15].............................................................................................................................7

Figure 1.6 Basic gas turbine with steam injection (STIG) plant configuration

[16]…………………………………………………………………………………………8

Figure 1.7 Schematic of Oxy-fuel combustion with CO2 capture ………………….10

Figure 1.8 Matiant cycle burning methane with oxygen and with CO2 removal [16]

……………………………………………………………………………………………12

Figure 2.1 Work for compression CO2 from atmospheric pressure to a given end

pressure. [26] …………………………………………………………………………...23

Figure 3.1 Schematic diagram of configuration 1, 2 and 3 …………………………29

Figure 3.2 Schematic diagram of configuration 4, 5 and 6 …………………………31

Figure 3.3 Configuration of Heat Recovery Steam Generator (HRSG) …………..46

Figure 4.1 Effect of pressure ratio on combined cycle thermal efficiency ………..52

Figure 4.2 Effect of pressure ratio on combined cycle exergy efficiency …………53

Figure 4.3 Effect of pressure ratio on the topping cycle work output with 5% steam

injection ………………………………………………………………………………….54

Figure 4.4 Effect of pressure ratio on the steam cycle work output with 5 % steam

injection ………………………………………………………………………………….55

Figure 4.5 Effect of pressure ratio on the combined cycle work output with 5 %

steam injection ………………………………………………………………………….55

Figure 4.6 Effect of steam injection in CC1 and CC2 on Carbon Dioxide Emission

…………………………………………………………………………………………....57

ix

Figure 4.7 Effect of ambient air temperature on cycle work output with steam for

configuration 2 ………………………………………………………………………….58

Figure 4.8 Effect of ambient air temperature on cycle work output for configuration

3 ………………………………………………………………………………………….59

Figure 4.9 Effect on work output with steam injection for configuration 1, 2 and 3

……………………………………………………………………………………………62

Figure 4.10 Effect of various TIT on combined cycle work output for configuration 1

……………………………………………………………………………………………64

Figure 4.11 Effect of TIT on topping cycle work output for configuration 2 ……....64

Figure 4.12 Effect of TIT on topping cycle work output for configuration 3 ………65

Figure 4.13 Effect of TIT on steam cycle work output for configuration 2 ………..66

Figure 4.14 Effect of TIT on steam cycle work output for configuration 3 ………..66

Figure 4.15 Effect of TIT on combined cycle work output with fraction of steam

injection in CC1 ………………………………………………………………………...67

Figure 4.16 Effect of TIT on combined cycle work output with fraction of steam

injection in CC1 and CC2 …………………………………………...………………...67

Figure 4.17 Effect of Isentropic efficiency on Topping cycle work output with

fraction of steam injection in CC1 …………………………………..………………...68

Figure 4.18 Effect of Isentropic efficiency on bottom cycle work output with fraction

of steam injection in CC1 ……………………………………………………………...69

Figure 4.19 Effect of Isentropic efficiency on combined cycle work output with

fraction of steam injection in CC1 …………………………………..………………...69

Figure 4.20 Effect of steam injection on turbine inlet temperature with fraction of

flue gas recycle for configuration 5 …………………………………………………...71

Figure 4.21 Effect of steam injection on turbine inlet temperature with fraction of

flue gas recycle for configuration 6 …………………………………………………...72

Figure 4.22 Effect of steam injection on combined cycle thermal efficiency with

fraction of flue gas recycle for configuration 5 & 6 ………………..………………...72

x

Figure 4.23 Effect of pressure ratio on topping cycle work output for configurations

4, 5 and 6 …………………………………………….……………….………………...74

Figure 4.24 Effect of pressure ratio on steam cycle work output for configurations

4, 5 and 6 …………………………………………….……………….………………...74

Figure 4.25 Effect of pressure ratio on combined cycle work output for

configurations 4, 5 and 6 …………………………………………….………………...75

Figure 4.26 Effect of pressure ratio on combined cycle thermal efficiency for configurations 4, 5 and 6 …………………………………………….………………...75

Figure 4.27 Effect of pressure ratio on exergy efficiency for configurations 4, 5 and

6 ………………………………………………………………………..………………...76

Figure 4.28 Effect of pressure ratio on CO2 emission for configurations 4, 5 and 6

……..………………..................................................................................................76

Figure 4.29 Percentage of exergy destruction in each component as compared to

overall destruction in combined cycle with steam injection …….……………….....77

xi

List of tables

Table 1.1 - World primary energy supply and CO2 emission: shared by fuel in 2007

[1] ……….………………............................................................................................1

Table 1.2 - World sources of electricity generation by fuel type in 2009 …...………2

Table 3.1 Natural gas combined cycle power generation configurations description

……………………………………………………………………………………………27

Table 3.2 Ideal-gas specific heats of various exhaust gases [18] …………………37

Table 4.1 Operating parameters of combined cycle configuration for various

pressure ratio (configuration 1, 2 and 3) ……………………………………………..52

Table 4.2 Operating parameters of combined cycle configuration to determine CO2

emission (configuration 1, 2 and 3) …………………………………………………..56

Table 4.3 Operating parameters of combined cycle configuration for variable

ambient temperature (configuration 1, 2 and 3) …………………………………….57

Table 4.4 Fuel balance on CC1 and CC2 for configuration 1 ……………………..59

Table 4.5 Work output of combined cycle for configuration 3 ……………………..61

Table 4.6 Operating parameters of combined cycle configuration for various

turbine inlet temperature (configuration 1, 2 & 3) …………………………………..63

Table 4.7 Operating parameters of combined cycle configuration for variable

isentropic efficiency (configuration 1, 2 and 3) ………………………………….…..68

Table 4.8 Effect of flue gas recycle on turbine inlet temperature in combined cycle

system for configuration 4 ………………………………………………………….....70

Table 4.9 Combined cycle thermal efficiency validation with reference [31] ….....78

Table 4.10 Combined cycle thermal efficiency validation with reference [15]….…79

Table 4.11 Combined cycle component exergy destruction validation with

references [14] and [15] …………………..…………………………………………...80

Table 4.12 Combined cycle component exergy efficiency validation with reference

[15] ………………………………………………………………………………………80

xii

Nomenclature

Symbols

CO2 = Carbon Dioxide

CH4 = Methane Gas

Cp = Specific heat at constant pressure (kJ/kmol.K)

��𝑊 = Work (kW)

��𝐷 = Exergy destruction (kW)

𝑒𝑝ℎ = Physical exergy (kW)

𝑒𝑐ℎ = Chemical exergy (kJ/kmol)

h = Enthalpy (kJ/kg)

ℎ𝑓0 = Enthalpy of formation (kJ/kg)

ℎ = Enthalpy at state (kJ/kg)

ℎ0 = Enthalpy at ambient condition (kJ/kg)

𝐼 = Entropy generation (kJ/kmol.K)

kW = Kilo Watt

m3 = Meter Cube

Mi = Molecular weight (kg)

mr = Mass of reactant in combustion chamber (kg)

mp = Mass of product in combustion chamber (kg)

�� = mass flow rate (kg/s)

N2 = Nitrogen Gas

N = Nitrogen Mole

xiii

O2 = Oxygen Gas

O = Oxygen Mole

Pr = Pressure Ratio

P = Pressure (Bar)

𝑅𝑢 = Universal gas constant

��gen Entropy generation (kJ/kmol.K)

s = Entropy (kJ/kmol.K)

To = Ambient temperature (K)

�� = Work (kW)

��𝐺𝑇1 = Main gas turbine work output (kW)

��𝐺𝑇2 = Reheater gas turbine work output (kW)

��𝑆𝑇 = Steam turbine work output (kW)

�� = Exergy rate (kW)

Xdestroyed = Exegy destroyed

Acronyms

ASU = Air Separation Unit

C1 = First stage compressor

C2 = Second stage compressor

CC = Combined Cycle

CC1 = Main Combustion Chamber

CC2 = Reheater Combustion Chamber

CCPP = Combined Cycle Power Plant

xiv

CCPP = Combined Cycle Power Plant

COND = Condenser

GW.h = Giga Watt Hour

GHG = Green House Gases

GT1 = Gas Turbine after main combustion chamber

GT2 = Gas Turbine after reheater combustion chamber HAT = Humid Air Turbine cycle

HP = High Pressure

HRSG = Heat Recovery Steam Generator

IC = Intercooler

IGCC = Integrated Gasification Combined Cycle

LP = Low Pressure

MW = Mega Watt

MT = Megatonnes

NG = Natural Gas

NOx = Generic term for nitric oxide NO or nitrogen oxide NO2

OTSG = Once Through Steam Generator

P = Pump

ST = Steam Turbine

STIG = Steam Injected Gas Turbine

Theoair = Theoretical air

xv

Greek Letters

𝛼 = Fuel in main combustion chamber (kg/s)

𝛽 = Fuel in reheater combustion chamber (kg/s)

𝜁 = Fraction of steam injection in main combustion chamber (%)

𝜔 = Fraction of steam injection in reheater combustion chamber (%)

𝜙 = Molar mass of oxygen (kg/kmol)

𝜃 = Molar mass of nitrogen (kg/kmol)

𝜆 = Molar mass of flue gas (kg/kmol)

η𝑐 = Compressor isentropic efficiency

η𝑔 = Gas turbine isentropic efficiency

η𝑠𝑡 = Steam turbine isentropic efficiency

η𝑝 = Pump isentropic efficiency

1

Chapter 1: Introduction

_________________________________________________________________

1.1 Energy scenario

The world’s dependence on fossil fuels for the satisfaction of primary energy needs

is at odds with growing atmospheric emissions of CO2 from the combustion of

hydrocarbons. Given their high energy density and availability, fossil fuels are likely

to continue to provide more than 80% of total world energy requirements for the

coming decades, with especially coal and natural gas asserting their positions in the

fuel mix by providing 38% and 30%, respectively of electricity demand in 2030 [1].

On a global basis, coal accounted for 26% of primary energy consumption in 2007,

oil for 34%, natural gas for 21%, nuclear for 5%, large hydropower for 5% and

renewable accounting for approximately 9%. Though coal represented only a

quarter of the world primary energy supply in 2007, it accounted for 42% of the global

CO2 emissions due to its heavy carbon content per unit of energy released. As

compared to gas, coal is on average nearly twice as emission intensive [1].

Table 1.1 - World primary energy supply and CO2 emission: shared by fuel in 2007 [1]

Oil Coal Gas Other*

Total Primary Energy

Supply (TPES) 34% 26% 21% 19%

CO2 emission 38% 42% 20% 0%

*other includes nuclear, hydro, geothermal, solar, tide, wind

2

Figure 1.1 Power generation comprises the largest source of CO2

emissions in 2007 [1]

The production of electricity in world (2009) was 20,053TWh. Sources of electricity

were fossil fuels 67%, renewable energy 16% (mainly hydroelectric, wind, solar and

biomass), and nuclear power 13%, and other sources were 3%. The majority of fossil

fuel usage for the generation of electricity was coal and gas.

Table 1.2 - World sources of electricity generation by fuel type in 2009

Coal Oil

Natural Nuclear Renewables other Total

Gas

Average electric power (GW)

943 127 490.7 311.6 375.1 64.8 2311

Proportion 41% 5% 21% 13% 16% 3% 100%

*Source: International Energy Agency (IEA)

Although the share of unconventional emerging technologies is expected to remain

small at less than 15 % by 2020, large changes are projected in the magnitude of

these generation technologies. Most notable is wind capacity, which is projected to

form 10% of total installed capacity by 2020. Other generation technologies such as

biomass, landfill gas, waste heat, solar and tidal also grow significantly. By 2020,

Electricity Generation,

41%

Transport, 23%

Industry, 20%

Agriculture, 6%

Other*, 10%

3

technologies such as carbon capture and storage (CCS) are expected to be used

more broadly to contain emissions from fossil fuel power generation [1].

Figure 1.2 Canada electricity generation by fuel types [1]

1.2 Natural gas

Canada has large amounts of remaining natural gas estimated at 12424x109

m3 (439 Tcf – Trillion cubic feet). Natural gas will continue to be relied upon to meet

increased electricity demand. Natural gas-fired generation is forecast to increase

during the period of the 2009 Reference Case Scenario by an additional 5 517 MW

of combined-cycle generation and 2 629 MW of combustion turbine/cogeneration

facilities. A decrease of 1 243 MW of steam turbine generation is also assumed

mainly in British Columbia (630 MW) and Alberta (221 MW), as older plants are

replaced by more efficient combined-cycle facilities. In the near term, investment

in combined-cycle generation is planned for Ontario as well as Newfoundland and

Labrador [1].

Natural gas-fired generation output is forecast to increase significantly from

50 809 GWh to 82 670 GWh in 2020, or from 8.4 % to 11.7 % of total generation.

4

In Ontario, a combination of 3 917 MW of combined-cycle gas and 1 337 MW of

combustion turbine/cogeneration facilities will be relied on to help meet demand

following the phase-out of coal-fired generation.

Figure 1.3 Electricity generations in Ontario by various energy sources [2]

Gas fired power sector is growing. Ontario government is showing support for gas

fired combined cycle power generation. Also natural gas prices support gas fired

power generation.

1.3 Carbon Dioxide Emission

Electricity generation using carbon based fuels is responsible for a large fraction

of carbon dioxide (CO2) emissions worldwide. Atmospheric CO2 concentration has

increased from 280 to 380 ppm by volume (2005); a 35% change since pre-

industrial time, largely due to carbon emissions from anthropogenic fossil fuel

burning and deforestation. The emission rate of carbon from fossil fuel (oil, coal

and gas) consumption is currently about 8 Gigaton Carbon per year, while the

deforestation rate is estimated to be 1.6 Gigaton per year. The cumulative fossil

fuel emission since 1800 is 330 Gigaton, but only about half of that remains in the

atmosphere; the remainder absorbed by carbon sinks in the ocean and on land [3-

4].

5

Jacob [5] describes the mean lifetime of CO2 in the atmosphere is much longer years

than previous estimations before being removed by ocean, photosynthesis or other

processes. Stabilising concentration of CO2 at any level would require large

reductions of global CO2 emissions from current level. The lower the chosen level

for stabilisation, the sooner the decline in global CO2 emissions would need to begin.

The supply of gas in 2007 was almost three times higher than in 1971 and its share

in emissions increased by five percentage points over that period [3]. Figure 1.3

shows the electricity generation by energy source in Ontario. Electricity production

from natural gas is gaining attention in Ontario. Also it is more feasible to implement

a CO2 capture system at the power generation stations due to its stable and constant

operation.

1.4 Combined cycle power plants

On account of the ever-increasing demand of electric power, several new

technologies have been developed during the last two decades. The thrust is mainly

in the direction of increasing the efficiency of generation and the capacity of

individual units and the entire power plants. New possibilities have been examined

for large base load power plants, and more suitable and reliable peak load plants.

In some countries gas turbine power plants are preferred on account of the

abundantly available fuel oil and natural gas. Recently large combined cycle power

plants have also become popular.

6

Figure 1.4 Combined cycle operation with Brayton and Rankine cycle

[18]

The gas turbine power plant has a main disadvantage of a lower thermal efficiency

and capacity compared to the hydro and steam turbine power plants. Attempts to

improve the efficiency and output power of the gas turbine power plants by

employing regeneration, intercooling and reheating result in significant pressure

losses on account of longer gas flow passages. This also increases the capital and

maintenance costs.

It has been found that a considerable amount of heat energy goes as a waste with

the exhaust of the gas turbine. This energy must be utilized. The complete use of

the energy available to a system is called the total energy approach. The objective

of this approach is to use all of the heat energy in a power system at the different

temperature levels at which it becomes available to produce work, or steam, or the

heating of air or water, thereby rejecting a minimum of energy waste. The best

approach is the use of combined cycles.

7

As shown in figure 1.5, there may be various combinations of the combined cycles

depending upon the place or country requirements. Combined cycle power plant is

a combination of an open cycle gas turbine and steam turbine. The exhaust of gas

turbine which has high oxygen content is used as the inlet gas to the steam

generator where the combustion of additional fuel takes place. This combination

allows nearer equality between the power outputs of the two units than is obtained

with the simple recuperative heat exchanger. For a given total power output the

energy input is reduced and the installed cost of gas turbine per unit of power output

is about one-fourth of that of steam turbine.

Figure 1.5 Combined cycle power plant with reheater combustion chamber

[15]

8

In other words, the combination cycles exhibit higher efficiency. The greater

disadvantages include the complexity of the plant, different fuel requirements and

possible loss of flexibility and reliability. The most recent technology in the field of

co-generation developed utilizes the gaseous fuel in the combustion chambers

produced by the gasification of low quality of coal. The system is efficient and the

cost of power production per kW is less.

1.5 Gas Turbine with Steam Injection (STIG)

Figure 1.6 shows basic configuration of gas turbine power generation system with

steam injection. Cheng [22] proposed a cycle in 1978 in which a fraction of the steam

produced from once through steam generator (OTSG) or heat recovery steam

generator (HRSG) injected into the gas turbine combustion chamber with fuel and

air.

C = Air Compressor, B = Combustion Chamber, T = Gas Turbine, P = Water Pump

Figure 1.6 Basic gas turbine with steam injection (STIG) plant configuration

[16]

9

Gallo [23] has compared simple gas turbine cycle with humid air turbine cycle (HAT),

gas turbine with steam injection (STIG) and other combinations. Simple gas turbine

cycles performs with best efficiency at TIT = 1573 K and a pressure ratio of 30.

Inclusion of intercooler between low pressure and high pressure compressors

strongly increases the performance of the cycle. In the STIG cycle, he noticed that

the pressure of the steam should be equal to the pressure in the combustion

chamber because unrestrained expansion of the steam until combustion chamber

pressure generates only irreversibility. The injected steam amount is 1% - 10% of

the air mass flow. STIG has higher efficiency than combine cycle without steam

injection. HAT cycle has the highest efficiency and lowest NOx formation due to

lower flame temperature. STIG cycle gives better performance at lower pressure

ratio. HAT, STIG and combined cycle gas turbine (CCGT) have higher water need

and water must be distilled at high quality. Such plants are suitable for high power

production.

1.6 Carbon capture mechanism and advances

In order to reduce the CO2 emissions from natural-gas (NG) based power-generation

plants, three different promising approaches have emerged [20 - 24].

(1) Post Combustion - Separation of CO2 from the exhaust gas of a standard gas-

turbine combined cycle (CC), using chemical absorption by amine solutions. This

approach has been widely treated in the literature and can be applied to the existing,

conventional plants. With monoethanolamine, alcanolamine, NaOH, membrane

separation, refrigeration and others.

(2) Oxy-fuel Combustion – Combined cycle with a close-to-stoichiometric

combustion with high-purity oxygen from an air-separation plant. As the combustion

10

products are CO2 and water vapour, in principle, CO2 can be captured simply by

condensing water from the flue gas.

(3) Precombustion - Decarbonization and CO2 capture, where the carbon of the NG

is removed prior to combustion and the fuel heating value is transferred to hydrogen

by reforming or partially combusted to CO and shifted towards H2 and CO2, with

subsequent separation. The resulting hydrogen can be burned in any cycle.

In the cases where the oxygen purity is below 99.5%, a low temperature inert gas

removal system is necessary, and the refrigeration required for its operation is

obtained by recycling CO2 gas around one of the stages of the nine stage centrifugal

CO2 compression train. The inerts are removed by phase separation at a

temperature which is fixed at an approach to the triple point temperature of CO2 (-

56.6°C). Operating as cold as possible will minimise loss of CO2 with the inert waste

gas but it will consume more power. For the design case, a reduction in temperature

of 2°C at the cold end results in an increase of 0.27 tonne/hr of CO2 captured with

an extra 100 kW power consumption. This is equivalent to 370 kWh/tonne of CO2,

giving a marginal cost for the extra CO2 captured of $16.1/tonne CO2 (at an

electricity cost of 4.35 c/kWh) compared with an overall cost of $20.9/tonne of CO2,

making the selection of the lowest possible operating temperature the best option

[25].

It should be noticed that CO2 is not completely recovered in power cycles with H2O

condensers due to solubility of CO2 in water. However, the solubility of CO2 in the

specific systems investigated is maximum 1% at 1 bar, which corresponds to about

4 g/kWh. This value is acceptable compared to a conventional Combined Cycle,

which emits about 400 g/kWh of CO2. The solubility is even lower at lower pressure.

The compression of CO2 takes place with intercooling, so that the compression work

11

is significantly less than the expansion work for a given equal pressure ratio in

expansion/compression [28].

1.7 Oxy-fuel combustion

With oxy-fuel combustion, oxygen is used in the form of a high-purity oxidant stream.

This enables combustion in a nitrogen-depleted atmosphere. This process results

in the production of a flue gas that is highly concentrated in CO2, thus simplifying the

CO2 capture process. An oxygen production plant is necessary.

Figure 1.7 Schematic of Oxy-fuel combustion with CO2 capture

Cooling the exhaust below the dew point enables the water to condense and the

resulting CO2 stream is obtained without the need for chemical absorption. A

commercial-scale gas-fired oxy-combustion power plant requires hundreds of tons

of oxygen each day. Currently, cryogenic distillation is the only commercially viable

technology that will produce such large quantities of O2. Other air separation

technologies like pressure swing adsorption, vacuum swing adsorption, and

polymeric membranes cannot economically produce such quantities.

Ceramic membranes (oxygen ion transport membranes) are not yet commercially

available for large-scale oxygen production, therefore making it difficult to compare

them to cryogenic distillation, both in terms of investment and performance [13].

Burning pure oxygen with methane can produce a significant high combustion

temperature about 3500° C. To bring the high temperature down for suitable use in

12

existing gas turbine, either parts of CO2 have to be recycled making CO2 as a

working fluid or water/steam has to be injected [17].

Horlock [16] described zero-emission “The Matiant Cycle” shown in figure 1.8.

Figure 1.8 Matiant cycle burning methane with oxygen and with CO2 removal

[16]

Matiant plant is more complex and ingenious version of the semi-closed cycle

burning fuel with oxygen. A stage of reheat and three stages of compression are

involved together with recuperator. CO2 and H2O are the working gases but both

the gases are removed through water separation, compression and liquefaction.

The multiple reheating and intercooling implies that such a cycle should attain high

efficiency. 55% of thermal efficiency is calculated at a maximum cycle pressure of

250 bar and combustion temperature of 1400° C [16].

13

1.8 Air separation unit (ASU)

Cryogenic ASU performances have improved tremendously over the last forty years.

It is estimated that power consumption has been cut almost in half, while distillation

column productivity (i.e., flow per square meter) has multiplied threefold. The

technology should continue to advance over the next decade, specifically through

targeted improvements in oxy‑combustion plants. Oxy-fuel combustion is chiefly

characterized by three elements: size (typically over 8 000 metric tons per day for

industrial-scale plants); low pressure (between 1.1 and 1.7 bar absolute); and

potentially low oxygen purity. Low oxygen purity would mean a value in the range of

85-98% O2 content compared to the typical 99.5-99.8% O2 content of high-purity

units. Using low purity O2 enables significant ASU power consumption savings [13].

The cycles for the production of low purity oxygen at 95% were developed in the

early 1990s, primarily for two applications: gasification (including IGCC) and oxygen

enrichment of blast furnace vent streams. These applications required the design of

plants that demonstrated specific separation energy around 200 kilowatt-hour per

metric ton (kWh/t) of pure O2. Separation energy is defined as the power required to

produce 1 metric ton of pure oxygen contained in a gaseous oxygen stream for a

given oxygen purity at an atmospheric pressure (101325 Pa) under ISO conditions

of 15°C and 60% relative humidity. Compressor driver efficiency (for electrical,

steam, or gas turbines), heat for regeneration of driers, and power consumption of

the cooling system are not considered in this definition [13].

Bollond (1992) calculated the energy consumption for oxygen production through

ASU and compression at 35 bar to 0.42 kWh/kg O2 [27] and 906 kJ/kg oxygen (0.25

kWh/kg) at atmosphere pressure.

14

1.9 Exergy

The energy neither created nor destroyed. Constant energy is always conserved.

What is not conserved is exergy, which is the useful work potential of the energy.

Once the exergy is wasted, it can never be recovered. When we use energy, we

are not destroying any energy; we are merely converting it to a less useful form,

a form of less exergy.

The useful work potential of a system at the specified state is called exergy. Exergy

is a property and is associated with the state of the system and the environment.

A system that is in equilibrium with its surroundings has zero exergy and is said

to be at the dead state. The exergy of heat supplied by thermal energy reservoirs

is equivalent to the work output of a Carnot heat engine operating between the

reservoir and the environment.

Reversible work (Wrev) is defined as the maximum amount of useful work that

can be produced (or the minimum work that needs to be supplied) as a system

undergoes a process between the specified initial and final states. This is the

useful work output (or input) obtained when the process between the initial and

final states is executed in a totally reversible manner. The difference between the

reversible work (Wrev) and the useful work (Wu) is due to the irreversibilities

present during the process and is called the irreversibility. It is equivalent to the

exergy destroyed. For a totally reversible process, the useful and reversible work

terms are identical and thus exergy destruction is zero. Exergy destroyed

represents the lost work potential and is also called the wasted work or lost work.

The second-law efficiency is a measure of the performance of a device relative to

the performance under reversible conditions for the same end states. Exergy can

be transferred by heat, work, and mass flow [18 -19].

15

1.9.1 Exergy destruction

Irreversibilities such as friction, mixing, chemical reactions, heat transfer through a

finite temperature difference, unrestrained expansion, nonquasiequilibrium

compression or expansion always generate entropy, and entropy always destroys

exergy. The exergy destroyed is proportional to the entropy generated and is

expressed as

Xdestroyed = To Sgen

This equation is applicable to any system undergoing any kind of process since any

system and its surroundings can be enclosed by a sufficiently large arbitrary

boundary across which there is no heat, work, and mass transfer, and thus any

system and its surroundings constitute an isolated system. No actual process is

truly reversible, and thus some exergy is destroyed during a process. The more

irreversible a process is, the larger the exergy destruction during that process. No

exergy is destroyed during a reversible process [18 - 19].

1.10 Objective of the thesis work

The objectives of the present thesis work are described below.

i) To develop and propose different gas turbine combine cycle power

generation configuration systems and to conduct thermodynamic

analysis of gas turbine combined cycle (GTCC) configuration with steam

injection

ii) Conduct parametric study of various gas turbine combined cycle power

generation configuration systems with varying operating conditions such

16

as turbine inlet temperature (TIT), ambient temperature, pressure ratio

and steam injection ratio.

iii) Steam injection effect on flue gas recycle, thermal efficiency, pressure

ratio for oxy-fuel combustion system

iv) Emission characteristics of the various power generation configuration

systems with steam injection

17

Chapter 2: Literature review: Recent work

_________________________________________________________________________

Present study is focused on natural gas fired combined cycle system for electricity

generation. In this system, topping cycle is brayton cycle and bottoming cycle is

rankin cycle. Primary focus is given to steam injection in combustion chamber and

reheater combustion chambers, steam injection effect on thermodynamic parameter

of overall cycle overall cycle and oxy-fuel cycle.

Franco and Casarosa [8] have studied on some perspectives for increasing the

efficiency of combined cycle power plants. The paper proposes an analysis of some

possibilities to increase the combined cycle plant efficiency to values higher than the

60% without resorting to a new gas turbine technology. This study reveals that the

optimization of heat recovery steam generator (HRSG) with the use of parallel

sections and of limit subcritical conditions (up to 220 bar) is the key elements to

obtain this result. They found that HRSG optimization is sufficient to obtain combined

cycle plant efficiencies of the order of 60% while, joining HRSG optimization with the

use of gas turbine reheat (postcombustion) and gas to gas recuperation can lead

the efficiency of the whole plant to the limit value of 65%. Results of this study are

proposed with reference to a turbine inlet temperature of 1500 K, corresponding to

those of usual commercial D-F series gas turbine [8].

Dincer [9] has studied the role of exergy in energy policy making. This paper deals

with the utilization of exergy as an efficient tool for energy policy making applications

since exergy is a measure of quantity and quality of the exergy sources unlike

energy which is only about the quantity. In addition, energy and exergy concepts are

evaluated for various actual process, and the role of exergy is discussed for energy

and environment policymaking activities from several key perspectives, e.g., quality,

energy conservation environment, economy, and sustainable developments. The

results of this study exhibit that the potential usefulness, exergy, in addressing and

18

solving environment problems as well as attaining sustainable development is

crucial. The researcher has concluded some crucial remarks on exergy from this

study as:

It is an effective method using the conservation of mass and conservation of

energy principles together with the second law of thermodynamics for the

design and analysis of energy systems.

It is the best primary tool in addressing the impact of energy resource use on

the environment and a key component in obtaining sustainable development.

It is an efficient technique revealing whether or not and by how much it is

possible to design more efficient energy systems by reducing the

inefficiencies in existing systems and distinguishing the quality between

energy resources.

Leo et al. [10] has studied gas turbine turbocharged by a steam turbine: a gas turbine

solution increasing combined power plant efficiency and power. In this paper a new

design of a combined-cycle gas turbine power plant (CCGT) with sequential

combustion that increases efficiency and power output in relation to conventional

CCGT plants is studied. This innovative proposal consists fundamentally in using all

the power of the steam turbine to turbocharge the gas turbine. A computer program

has been developed in this study to carry out calculations and to evaluate

performance over a wide range of operating conditions. The obtained results are

compared with those of combined cycles where the gas turbines are not

turbocharged and the gas and the steam turbines have independent power exits.

The result shows that combined cycle efficiency has been increased from 58.14 %

to 58.24 % with the pressure ratio of 30. [10]

Sanjay et al. [11] have performed research work on energy and exergy analysis of

steam cooled reheat gas steam combined cycle. This research paper deals with

parametric energy and exergy analysis of reheat gas–steam combined cycle using

19

closed-loop-steam-cooling. They have compared the blade cooling techniques and

found that closed-loop-steam-cooling to be superior to air-film cooling. The reheat

gas–steam combined cycle plant with closed-loop-steam-cooling exhibits enhanced

thermal efficiency (around 62%) and plant specific work as compared to basic

steam–gas combined cycle with air-film cooling as well as closed-loop-steam

cooling. Further, with closed-loop-steam-cooling, the plant efficiency, reaches an

optimum value in higher range of compressor pressure ratio as compared to that in

film air cooling. They have also concluded that:

Reheat pressure is an important design parameter and its optimum value

gives maximum plant efficiency.

Air-film-cooling offers more exergy loss in compressor as compared to

closed-loop-steam cooling while gas turbine exergy loss is less in air-film-

cooling. Higher value of exergy in gas turbine is exhibited for closed-loop-

steam cooling, while higher value of steam turbine exergy is that for air-film-

cooling. Lower values of exergy losses are observed in compressor, gas

turbine, heat-recovery-steam-generator, stack, steam turbine, condenser

boiler-feed-pump and deaerator for closed-loop-steam cooling.[11]

Khaliq and Kaushik [12] have studied Second-law based thermodynamic analysis

of Brayton/Rankine combined power cycle with reheat. The aim of the present paper

is to use the second-law approach for the thermodynamic analysis of the reheat

combined Brayton/Rankine power cycle. Expressions involving the variables for

specific power-output, thermal efficiency, exergy destruction in components of the

combined cycle, second-law efficiency of each process of the gas-turbine cycle, and

second law efficiency of the steam power cycle have been derived. n this paper the

effects of pressure ratio, cycle temperature ratio, number of reheats and cycle

pressure-drop on the combined cycle performance parameters have been

investigated. The results of this study showed that the exergy destruction in the

20

combustion chamber represents over 50% of the total exergy destruction in the

overall cycle. The combined cycle efficiency and its power output were maximized

at an intermediate pressure-ratio of 32, and increased sharply up to two reheat-

stages and more slowly thereafter. Their main findings are:

The second-law efficiency of the adiabatic compressor increases with

pressure ratio because the absolute values of the work input and exergy

increase are both larger and the magnitude of exergy destruction in the

adiabatic compressor increases with the increase in pressure ratio

The first-law efficiency of the adiabatic turbine increases with the increase in

pressure ratio. The second-law efficiency decreases with the pressure ratio,

but increases with the cycle temperature ratio since a greater proportion of

the available work lost at the higher temperature may be recovered. The

exergy destruction in the reheat turbine increases with the pressure ratio.

At low pressure ratio, then the gas-turbine cycle and combined-cycle

efficiencies and their specific work-outputs drop, whereas the steam cycle

work-output increases due to the high gas-turbine exhaust temperature. At

an intermediate pressure-ratio, both the efficiency and specific work peak. At

high pressure ratio, the compressor and turbine works increase but their

difference, the net gas-turbine work output drops. The absolute magnitude of

exergy destroyed in both compressor and turbine increases as the logarithm

of pressure ratio. The exergy lost in the reheat turbine also increases due to

the lower mean temperature of reheat. The steam-turbine cycle output

suffers with the lower exhaust-gas temperature [12]

Butcher and Reddy [13] have studied Second law analysis of a waste heat recovery

based power generation system. In this paper the performance of a waste heat

recovery power generation system based on second law analysis is investigated for

various operating conditions. The temperature profiles across the heat recovery

21

steam generator (HRSG), network output, second law efficiency and entropy

generation number are simulated for various operating conditions. The variation in

specific heat with exhaust gas composition and temperature are accounted in the

analysis and results. The effect of pinch point on the performance of HRSG and on

entropy generation rate and second law efficiency are also investigated. The

researchers found that the second law efficiency of the HRSG and power generation

system decreases with increasing pinch point. The first and second law efficiency of

the power generation system varies with exhaust gas composition and with oxygen

content in the gas. The results contribute further information on the role of gas

composition, specific heat and pinch point influence on the performance of a waste

heat recovery based power generation system based on first and second law of

thermodynamics [13].

Ameri et al. [14] have studied the exergy analysis of a 420MW combined cycle power

plant. Their objective is to evaluate irreversibility of each part of Neka CCPP using

the exergy analysis. The results show that the combustion chamber, gas turbine,

duct burner and heat recovery steam generator (HRSG) are the main sources of

irreversibility representing more than 83% of the overall exergy losses. The results

show that the greatest exergy loss in the gas turbine occurs in the combustion

chamber due to its high irreversibility. As the second major exergy loss is in HRSG,

the optimization of HRSG has an important role in reducing the exergy loss of total

combined cycle [14].

Srinivas et al. [15] performed parametric simulation of steam injected gas turbine

combined cycle with the dual pressure heat-recovery steam generator. Effect of

operating variables such as low-pressure (LP) steam temperature ratio, steam

reheat pressure ratio, steam turbine inlet pressure, gas cycle pressure ratio and

combustion chamber temperature on the efficiency of the combined cycle has been

investigated. Exergy efficiency of the cycle is compared with and without the steam

22

injection with respect to the studied parameters. Maximum mass ratio of steam

injection to fuel has been examined as 6 kg/kg fuel with the complete combustion of

the fuel due to excess air supply in the combustion chamber and gas reheater. LP

temperature ratio is identified as a dominant parameter having impact on the

efficiency of the combined cycle as the steam is injected at this pressure. The results

showed that the major exergetic loss in the combustion chamber decreased with the

steam injection [15].

Bolland et al. [26] studied various configurations to capture CO2 capture options for

natural gas fired combined cycle plants. For the case of burning methane with pure

oxygen, he states that a standard gas turbine cannot be used for the purpose of a

stoichiometric combustion with O2 supplied from an air separation unit because the

optimal gas turbine pressure ratio is significantly higher than for gas turbines

operating with air. The pressure ratio was set to 35 bar for the oxy-fuel combustion

instead of the more typical 14–18 bar, which is typically found in existing as turbines.

With the same technology level as for an existing 250–300 MW class gas turbine

that would give a CC efficiency of 58%. The efficiency of the modified CC (with ‘free’

O2) was calculated to 61–62%, depending upon the temperature of supplied O2

(200–500 ˚C) [26].

The efficiency penalty for producing/compressing O2 is nearly 12%. Some of these

losses are recovered (the difference between 61.5% and 58%) because the O2 that

is fed to the gas turbine, at an elevated temperature and pressure, brings some

exergy into the gas turbine. The efficiency penalty for O2 production seems to be

significantly higher than for the capture of CO2 from exhaust gases (post combustion

CO2 recovery). Cryogenic separation of O2 from air is burdened with larger

irreversibility (pressure drop, heat transfer losses) than the absorption process for

capturing CO2 from the exhaust gas.

23

Figure 2.1 Work for compression CO2 from atmospheric pressure to a given

end pressure. [26]

CO2 recovered from the oxy-fuel combustion has to be compressed for subsequent

storage or further use. Figure 2.1 shows the compression with 3 intercoolers (15° C)

and compressor adiabatic efficiency ranging from 75% (high-pressure) to 85% (low-

pressure) [26].

Shyam et al. [30] studied effect of steam injection on regenerative gas turbine

system. They found 1% steam injection in regenerative cycle increases the work

output by 8.57 MW while 2% steam injection in the same configuration increases

work output by 19.05 MW. Also overall combined cycle thermal efficiency without

steam injection is 36.99 % compared to 40.44 % and 45.05% when 1% and 2%

steam injection respectively. Highest exergy destruction found in combustion

chamber followed by heat recovery steam generator, gas turbines, steam turbines

and compressors. Under the ambient temperature of 25° C, the benefit of adding the

STIG feature can substantially improve the power output from the 30 MW to 38.57

MW and power generation efficiency by 4.4%. The maximum power that can be

reached by the system with both IAC and STIG features is 49.05 MW for steam

24

injection pressure ratio at 0.2. Although the steam injection will increase the total

exergy losses, the exergy loss per MW output is much smaller than that of

regenerative cycle. It also reveals that the degree of energy wasting and thermal

pollution can be reduced through retrofitting [30].

S. Kumar [31] studied effect of gas turbine inlet temperatures on combined cycle

system. It is observed that effect of increasing turbine inlet temperature produces

exhaust gas temperature and its enthalpy. The net specific work of topping cycle is

higher than that of bottoming cycle at each turbine inlet temperature. The net specific

work and efficiency of combined cycle also increases with increasing turbine inlet

temperature. An increase in maximum steam temperature results into slight increase

in net specific work but a significant increase in combined cycle efficiency. The best

cycle performance is seen at a turbine inlet temperature of 1,900 K for maximum

steam temperature of 570° C, which gives the cycle efficiency of 60.9 % with net

specific work of 909 kJ/kg. The combined cycle net specific work decreases with

increasing pressure ratio and decreasing maximum steam generation temperature.

Both the cycle efficiency and net specific work increase with turbine inlet

temperature (TIT) for each pressure ratio and at any TIT there exists an optimum

pressure ratio for cycle efficiency and specific work [31].

Sanjay at al [32] studied performance evaluation of gas-steam combined cycle

having transpiration cooled gas turbine. The found that when TIT increases,

temperature of exhaust from gas turbine increases and also its enthalpy increases.

Since the maximum steam generation temperature in HRSG is fixed hence the

difference between the exhaust gas temperature and steam generation temperature

increases with increasing TIT. This increase in temperature difference augments the

heat recovery in HRSG, producing more steam in bottoming cycle. For an increase

of TIT from 1600 K to 1700 K the exhaust gas temperature increases from 846 K to

907 K and heat recovery percentage increases by 6.4% and an increase of 2.37%

25

in combined cycle efficiency. Overall cycle efficiency is 57% at 1600 K TIT and

increase to 61.3% at 1800 K TIT [32].

Alok at al. [33] performed parametric study of the effect of compressor pressure

ratio, compressor inlet temperature, turbine inlet temperature. Research shows clear

influence of turbine inlet temperature on plant work output. As the turbine inlet

temperature increases, it increase plant work output. Plant thermal efficiency

reached at 57% when TIT is 1850 K. Pressure ratio range of 16 to 28 studies and

compared with overall plant efficiency. It is concluded that higher pressure ratio

gives higher efficiency at particular turbine inlet temperature. Optimum pressure

ratio found to be 24 when TIT is 1850 K, while there is a decline in efficiency when

pressure ratio is higher than 24. They also concluded best ambient condition is when

ambient temperature is high and ambient relative humidity is low [33].

Sven at al. [34] studied oxyfuel combustion combined cycle where natural gas is

burned with pure oxygen and carbon dioxide is used to bring turbine inlet

temperature down to normal operating condition of turbine blades. They found that

pressure ratio of the oxyfuel gas turbine is much higher (about 40) compared to

conventional combined cycle gas turbine (about 18) due to relative low specific heat

ratio for carbon dioxide. Heat recovery steam generator does not differ from the one

used in conventional combined cycle system. Overall thermal efficiency is 63%

when not accounting the energy used for air separator unit (ASU) to produce pure

oxygen. To produce 106.6 MW of power, it consumes about 16.65 MW for ASU

accounting 15.7 % of turbine power output [34].

26

Chapter 3: Combine cycle power generation system configurations and

methodology

_________________________________________________________________

3.1 Natural gas combined cycle configuration description

Natural gas fired combined power generation systems are gaining popularity due to

their high combustion efficiency and reduced emission. Literature review shows that

overall efficiency can be increased greatly by various configurations on topping gas

cycle and bottoming steam cycle. In the present work natural gas combined system

is studied where natural gas is oxidized with air in primary and reheater combustion

chambers. Steam is extracted from the steam turbine at various percentage of air

mass flow rate and injected in either primary, reheater, or both combustion

chambers. Thermodynamic analysis is performed to understand energy and exergy

of the combined cycle systems. A similar study carried out with oxy-fuel combustion

where natural gas is oxidized with pure oxygen instead of air.

The present study is focused on combined cycle power generation configurations

described in table 3.1, where the topping cycle consists of an air compressor (C1)

followed by an intercooler (IC). Air is further compressed in second air compressor

(C2). Compressed air is burned with methane in a combustion chamber (CC1). In

case of configurations 2, 3, 5 and 6 described in table 3.1, a fraction is steam (ζ)

injected in the first combustion chamber. Products of first combustion chamber

enters main gas turbine (GT1) and produces work output. Exhaust gas from main

gas turbine further burned with methane in second combustion chamber (CC2). In

case of configuration 3 and 6, a fraction of steam ω is injected in reheater

combustion chamber. Product of reheater combustion chamber enters reheater gas

turbine (GT2) to produce work output. Exhaust gas from reheater gas turbine enters

heat recovery steam generator which produces steam in the bottoming cycle. Steam

enters steam turbine (ST) and produces work output. Fraction of steam (ζ and ω) is

27

taken out at particular pressure and used in topping cycle for power augmentation.

Water vapor from steam turbine condensed in condenser and recirculates in the

bottoming cycle through pump. Various configurations are studied as described in

table 3.1. It is assumed that system is operating at a steady state steady flow

conditions.

Table 3.1 Natural gas combined cycle power generation configurations

description

Configuration

#

Description of Study

1 Combustion with air and methane without steam injection

2 Combustion of air and methane with steam injection in main

combustion chamber

3 Combustion of air and methane with steam injection in main

combustion chamber and reheater combustion chamber

4 Combustion of oxygen and methane without steam injection

5 Combustion of oxygen and methane with steam injection in main

combustion chamber

6 Combustion of oxygen and methane with steam injection in main

combustion chamber and reheater combustion chamber

28

3.1.1 Natural gas combined cycle power generation configuration 1

Figure 3.1 shows the generalized schematic diagram of configuration 1, 2 and 3. Air

is compressed from ambient condition into air compressor (C1) and (C2). An

intercooler (IC) is used between C1 and C2 to bring the air temperature down and

to reduce overall compressor work. Compressed air (𝜙O2 + (𝜃*3.76) N2) enters

combustion chamber 1 (CC1) and burned with fuel methane (𝛼) at constant

pressure. Combusted gas has high thermal energy and expanded partially into gas

turbine 1 (GT1) to obtain shaft work (��𝐺𝑇1). Partially expanded gas enters

combustion chamber 2 (CC2) and burned with fuel methane (𝛽) to elevate the

thermal energy of the gas. Gas coming out of CC2 is expanded in gas turbine 2

(GT2) to obtain shaft work (��𝐺𝑇2).

Gas coming out of GT2 is passed through Heat Recovery steam Generator (HRSG)

to make steam which passes through Steam Turbine (ST) to obtain shaft work (��𝑆𝑇).

Expanded steam in ST is brought to condenser (COND) to extract latent heat and

change state to saturated water. Saturated water is passed through pump (P) to

increase pressure and passed through HRSG to complete the Rankin cycle.

29

Figure 3.1 Schematic diagram of configuration 1, 2 and 3

3.1.2 Natural gas combined cycle power generation configuration 2

Configuration 2 has basic schematic same as configuration 1 except, a fraction of

steam (𝜻) is extracted from ST and injected inside combustion chamber CC1 with

air and fuel. ζ is a percentage of air mass flow which enters combustion chamber

CC1 at state 4. ζ is taken out from steam turbine at the stage where steam pressure

is 5% higher than pressure of combustion chamber 1.

3.1.3 Natural gas combined cycle power generation configuration 3

As an extension of configuration 2, fraction of steam (𝜔) is injected inside of reheater

combustion chamber CC2. Steam is taken out of steam turbine at 5% higher

pressure than pressure present inside combustion chamber 2

30

3.1.4 Natural gas combined cycle power generation configuration 4 (oxy-

fuel cycle)

Figure 3.2 shows the generalized schematic diagram of configurations 4, 5 and 6.

In configuration 4, pure oxygen (𝜙) obtained from air separator unit (ASU) is used

as an oxidizing agent and burned in CC1 and CC2 with main combustion chamber

fuel 𝛼 and reheater combustion chamber fuel supply 𝛽 respectively instead of using

air as an oxidizing agent used in configuration 1, 2 and 3. Expanded gas after GT2

is passed through HRSG and steam is generated to operate bottoming Rankin cycle.

Exhaust gas after HRSG contains only CO2 and water vapour. Water vapour is

condensed through Water Separator (WS) and removed from exhaust gas.

Combustion of methane and pure oxygen produces very high flame temperature

[17] which is not suitable for turbine operation at this stage. A fraction of CO2, defined

as (𝜆) is compressed in three compressors (C3, C4, C5) coupled with an intercooler

between each compression. Highly compressed carbon dioxide with mole fraction

𝜆CO2 is passed through condenser until 𝜆CO2 changes it’s phase and coverts to

liquid phase. Liquid 𝜆CO2 is removed from the cycle for sequestration. Remaining

part of CO2 defined as (1- 𝜆) is compressed in C1 and C2 and recycled back to CC1

to bring flame temperature down to suitable operative condition.

3.1.5 Natural gas combined cycle power generation configuration 5 (oxy-

fuel cycle)

Basic schematic of configuration 5 is similar to configuration 4 except, a fraction of

steam (𝜻) is extracted from ST and injected inside CC1 with pure oxygen (𝜙) and

fuel (𝛼). CC2 is not injected with steam. Steam is taken from steam turbine at 5%

higher pressure than pressure present inside combustion chamber 1.

31

3.1.6 Natural gas combined cycle power generation configuration 6 (oxy-

fuel cycle)

As an extension of configuration 5, additional steam is taken from steam turbine (𝝎)

and injected inside of combustion chamber 2. Pressure at which (𝝎) is extracted

from steam turbine is 5% higher than pressure present at combustion chamber 2.

Figure 3.2 Schematic diagram of configuration 4, 5 and 6

3.2 Thermodynamic Analysis

Energy technologies are normally examined using energy analysis. A better

understanding is attained when a more complete thermodynamic view is taken,

which uses the second law of thermodynamics in conjunction with energy analysis,

via the exergy method. Through an exergy analysis, the efficiencies of processes

32

and devices are evaluated and the locations and sources of major inefficiencies are

identified. [29]

3.3 Methodology

In this chapter, each component in the combined cycle configurations is explained

for energy and exergy analysis. Each component is described for all configuration

described in table 3.1.

3.3.1 Gas compressors (C1) & (C2) and intercooler (IC)

For configuration 1, 2 and 3, air [𝜙O2 + (𝜃*3.76) N2] and for configuration 4, 5, and

6, mixture of oxygen and carbon dioxide [𝜙O2+ (1- 𝜆) CO2] enters in C1 (state 1) at

ambient conditions (P0, T0). Gases are compressed to the pressure P2 and enter

intercooler (IC). The work input to a compressor can be reduced by using multistage

compression with intercooling. For maximum savings from the work input, the

pressure ratio across each stage of the compressor must be the same. The main

purpose of the intercooler between C1 and C2 is to reduce overall compression work

input by reducing the heat of gases. The pressure ratio (Pr) represents the pressure

difference between state 4 and state 1.

Pr =𝑃4

𝑃1

(3.1)

The optimum intermediate pressure for compression is calculated as given below.

P2,3 = √P1 x P4 (3.2)

33

Since the relative pressure (P2) has been determined, the absolute entropy (s2s),

ideal enthalpy (h2s) and intercooler inlet temperature (T2s) can be interpolated from

the standard air tables or EES. In this study, Engineering Equation Solver (EES)

program has been used to determine all the state properties and related

calculations. The actual enthalpy at the first compressor outlet (state 2) is calculated

considering compressor isentropic efficiency (η𝑐)

h2 = h2s− h1

η𝑐 + h1

(3.3)

Heat is removed between states (2) and (3) through an intercooler. In an ideal case,

the temperature of the compressed gas as it leaves the intercooler (state 3) is the

same as the temperature at the inlet of C1 but this study, but a fair assumption of

temperature drop is assumed. Enthalpy (h3) and entropy (s3) are estimated at T3 and

P3. Gas properties at (state 4) are calculated similar manner to that discussed above

to calculate at (state 2).

h4 = h4s− h3

η𝑐 + h3

(3.4)

Atomic weight of Oxygen (O) = 16 (Kg/kmol) and Nitrogen (N) = 14 (Kg/kmol).

Therefore, molecular weight of air [𝜙O2 + (𝜃*3.76) N2] is [𝜙(2*16) + (𝜃*3.76)*(2*14)]

which is equivalent to [32𝜙+ 105.28𝜃] (Kg). Atomic weight of Carbon (C) = 12

(Kg/kmol), therefore molecular weight of [𝜙O2+ (1- 𝜆) CO2] is [32𝜙+ 44(1- 𝜆)] (Kg).

Mass flow rate for the oxidizing gas can be described as equation 3.5.

��gas = ��1 = ��2 = ��3= ��4 = 32𝜙+ 105.28𝜃 + 44(1 − 𝜆) (Kg/s)

(3.5)

34

By using energy balance equations for C1 and C2, we find and work input required

to compress gas at desired pressure ratio.

��C1 = ��gas (h2 – h1)

(3.6)

��C2 = ��gas (h4 – h3)

(3.7)

Physical exergy rate of C1 and C2, Exergy at the ambient state is considered to be

zero.

��2 = ��gas[(ℎ2 − ℎ0) − 𝑇0(𝑠2−𝑠0)] (3.8)

��3 = ��gas[(ℎ3 − ℎ0) − 𝑇0(𝑠3−𝑠0)] (3.9)

��4 = ��gas[(ℎ4 − ℎ0) − 𝑇0(𝑠4−𝑠0)]

(3.10)

Exergy destruction of C1 and C2 can be found from the exergy balance equations

described as,

𝐼��1 = (��1 − ��2) + ��C1 (3.11)

𝐼��2 = (��3 − ��4) + ��C2 (3.12)

Exergy destruction in intercooler,

𝐼��𝐶 = ��3 − ��2 (3.13)

Exergy Efficiency of C1 and C2,

35

𝜂𝑒𝑥𝐶1=

��2−��1

ẆC1

(3.14)

𝜂𝑒𝑥𝐶1=

𝑋4− 𝑋3

ẆC2

(3.15)

3.3.2 Primary and reheater combustion chambers CC1 and CC2

For all configurations, methane (CH4) is considered as fuel. Some assumptions are

made for the analysis. Purity of methane is 100%. Both combustion chambers are

working adiabatically, all the non-reacting gases are arbitrarily assigned as zero

thermomechanical enthalpy, entropy, and exergy at the condition of ambient

pressure and temperature regardless of their chemical composition. The entropy of

mixing different gaseous components is neglected. Kinetic and Potential energy

and related exergy is neglected. Combustion gases are ideal gases.

3.3.2.1 Primary combustion chamber CC1

The chemical reaction in combustion chamber 1 (CC1) is expressed by a chemical

equation 3.16. For configuration 1, ζ = 0 as there is no steam injection and λ = 1 as

there is no flue gas recycled in the system.

α CH4 + 𝜙 O2 + (𝜃*3.76) N2 + ζ H2O + (1-λ) CO2

(α+1-λ) CO2 + (2α+ζ) H2O + (𝜙-2α) O2 + (𝜃*3.76) N2 (3.16)

Mass of oxidizer = 𝑚𝑟𝑐𝑐1 = (32𝜙)O2+(105.28𝜃)N2+(18ζ)steam+[44(1-λ)]CO2 (Kg)

(3.17)

36

Mass of fuel = ��f1 = (16α)CH4 (Kg/s) (3.18)

Oxidizer Fuel ratio = 2𝜙 + 6.58𝜃 + 1.13ζ + 2.75(1−λ)

α (3.19)

Gas temperature (Tg1 ) combusted from CC1 can be found from eq. 3.20.

hr,5 = Cp5 Tg1 (3.20)

Total enthalpy of gas (h5) coming out after CC1 can be found from sum of molar

fractions of combustion products and the enthalpy of each gas component in gas

mixture at certain temperature dividing by the molar mass of the exhaust gases.

∑ 𝑦𝑖(ℎ𝑓𝑖

0+ ℎ𝑖

−𝑛

𝑖=0

ℎ0

)𝑟

hr,5 = (3.21) 𝑚𝑟𝑐𝑐1

Total mass of reactant (𝑚𝑟𝑐𝑐1) in CC1 is described in eq. 3.17. For the reactant

gases involved in CC1, eq. 3.21 can be written as,

ℎ𝑟,5 = {(𝛼 + 1 − 𝜆

𝑚𝑟𝑐𝑐1

) (ℎ𝑓0 + ℎ − ℎ0)}

𝐶𝑂25

+ {(𝜙 − 2𝛼

𝑚𝑟𝑐𝑐1

) (ℎ𝑓0 + ℎ − ℎ0)}

𝑂25

+ {(𝜃 ∗ 3.73

𝑚𝑟𝑐𝑐1

) (ℎ𝑓0 + ℎ − ℎ0)}

𝑁25

+ {(2𝛼 + 𝜁

𝑚𝑟𝑐𝑐1

) (ℎ𝑓0 + ℎ − ℎ0)}

𝐻2𝑂5

(3.22)

37

Enthalpy of formation is zero for O2 & N2 in eq. 3.22. The molar weight of the

combustion product (m5) (Kg) is calculated by multiplying the molars’ fraction by

molecular weight for each combustion product component shows in eq. 3.23

m5 = (α+1−λ)MCO2 + (2α+ζ)MH2O + (𝜙−2α)MO2 + (𝜃∗3.76)MN2

(α+1−λ)CO2 + (2α+ζ)H2O + (𝜙−2α)O2 + (𝜃∗3.76)N2

(3.23)

Specific heat (Cp) of an ideal gas mixture depends on temperature only and

independent of the pressure or the volume of the gas mixture. Cp at state 5, Cp5

(Kg/kmol.K) can be found from eq. 3.24.

Cp5 =

(α+1−λ)MCO2CpCO2 + (2α+ζ)MH2OCpH2O

+ (𝜙−2α)MO2CpO2 + (𝜃∗3.76)MN2CpN2

(α+1−λ)CO2 + (2α+ζ)H2O + (𝜙−2α)O2 + (𝜃∗3.76)N2

(3.24)

Specific 𝐶𝑝 of various gases at various temperatures can be found from equation

3.25. Where, T is the temperature at which the value of 𝐶𝑝 needs to be found and

value of constants a, b, c and d are given in table 3.2. [18]

𝐶𝑝 = 𝑎 + 𝑏𝑇 + 𝑐𝑇2 + 𝑑𝑇3 (3.25)

Table 3.2 Ideal-gas specific heats of various exhaust gases [18]

a b c d

N2 6.903 -0.02085 x 10-2 0.05957 x 10-5 -0.1176 x 10-9

O2 6.085 0.2017 x 10-2 -0.05275 x 10-5 0.05372 x 10-9

CO2 5.316 0.79361 x 10-2 -0.2581 x 10-5 0.3059 x 10-9

H2O 7.700 0.02552 x 10-2 0.07781 x 10-5 -0.1472 x 10-9

38

By trial and error, the TIT1 is set at a specific temperature, which can be achieved

by varying the amount of fuel (α) CH4 in the main combustion chamber and taking in

to account that this temperature should be in a reasonable range between 1200 K

and 1600 K.

Entropy generation in combustion chamber can be found from equation 3.26.

𝑠𝑔𝑒𝑛𝑐𝑐= 𝑠𝑝𝑟𝑜𝑑𝑢𝑐𝑡 − 𝑠𝑟𝑒𝑎𝑐𝑡𝑎𝑛𝑡 = ∑ 𝑁𝑝𝑠�� − ∑ 𝑁𝑟𝑠�� (3.26)

Absolute entropy values at standard pressures and temperature (P0, T0) can be

found from the air properties tables. Specific entropy at any condition (P, T) can be

found from eq. 3.27.

𝑠��(𝑇, 𝑃𝑖) = ��𝑖0(𝑇, 𝑃0) − 𝑅𝑢 ln

𝑦𝑖𝑃𝑚

𝑃0

(3.27)

where, Pi is the partial pressure, yi is the mole fraction of the component, and Pm is

the total pressure of the gas mixture. For the studied configurations, total mass of

product (𝑚𝑝𝑐𝑐1) and reactant (𝑚𝑟𝑐𝑐1

) in CC1 can be written separately as derived

from equation 3.16 as,

𝑚𝑟𝑐𝑐1= (α)CH4 + (𝜙)𝑂2 + (𝜃 ∗ 3.76)𝑁2 + (𝜁)𝑠𝑡𝑒𝑎𝑚 + (1 − 𝜆)𝐶𝑂2 (𝑘𝑚𝑜𝑙) (3.28)

𝑚𝑝𝑐𝑐1= (α + 1 − λ)CO2 + (2α + ζ)H2O + (𝜙 − 2α)O2

+ (𝜃 ∗ 3.76)N2 (kmol)

(3.29)

39

Entropy generation (𝑠𝑔𝑒𝑛𝑐𝑐1) of CC1 can be found from rewriting equation 3.26 and

3.27.

𝑠𝑟𝑒𝑎𝑐𝑡𝑎𝑛𝑡𝑐𝑐1= αsCH4 + 𝜙𝑠𝑂2,4 + (𝜃 ∗ 3.76)𝑠𝑁2,4 + 𝜁𝑠𝑆𝑇𝐸𝐴𝑀,4 + (1 − 𝜆)𝑠𝐶𝑂2,4

− 𝑅𝑢 [{⟨α𝑙𝑛 (αP4

𝑚𝑟𝑐𝑐1

)⟩𝐶𝐻4 + ⟨𝜙𝑙𝑛 (𝜙P4

𝑚𝑟𝑐𝑐1

)⟩𝑂2

+ ⟨(𝜃 ∗ 3.76)𝑙𝑛 ((𝜃 ∗ 3.76)P4

𝑚𝑟𝑐𝑐1

)⟩𝑁2 + ⟨𝜁𝑙𝑛 (𝜁P4

𝑚𝑟𝑐𝑐1

)⟩𝑠𝑡𝑒𝑎𝑚

+ ⟨(1 − 𝜆)𝑙𝑛 ((1 − 𝜆)P4

𝑚𝑟𝑐𝑐1

)⟩𝐶𝑂2}]

(3.30)

𝑠𝑝𝑟𝑜𝑑𝑢𝑐𝑡𝑐𝑐1= (α + 1 − λ)sCO2,5 + (2α + ζ)𝑠𝐻2𝑂,5 + (𝜙 − 2α)𝑠𝑂2,5

+ (𝜃 ∗ 3.76)𝑠𝑁2,5

− 𝑅𝑢 {(α + 1 − λ)𝑙𝑛 ((α + 1 − λ)P4

𝑚𝑝𝑐𝑐1

)

+ (2α + ζ)𝑙𝑛 ((2α + ζ)P4

𝑚𝑝𝑐𝑐1

) + (𝜙 − 2α)𝑙𝑛 ((𝜙 − 2α)P4

𝑚𝑝𝑐𝑐1

)

+ (𝜃 ∗ 3.76)𝑙𝑛 ((𝜃 ∗ 3.76)P4

𝑚𝑝𝑐𝑐1

)}

(3.31)

𝑠𝑔𝑒𝑛𝑐𝑐1= 𝑠𝑝𝑟𝑜𝑑𝑢𝑐𝑡𝑐𝑐1

− 𝑠𝑟𝑒𝑎𝑐𝑡𝑎𝑛𝑡𝑐𝑐1

(3.32)

Exergy destruction in CC1 can be found from equation 3.33.

𝐼d𝑐𝑐1= 𝑚5𝑇0 𝑠𝑔𝑒𝑛𝑐𝑐1

(3.33)

Exergy efficiency of CC1,

40

𝜂𝑒𝑥𝐶𝐶1= 1 −

𝐼𝑑𝑐𝑐1

𝑋5

(3.34)

where,

𝑋5 = 𝑚5[(ℎ5 − ℎ0) − 𝑇0(𝑠𝑝𝑟𝑜𝑑𝑢𝑐𝑡𝑐𝑐1−𝑠0) (3.35)

3.3.2.2 Reheater combustion chamber CC2

The combustion chamber reaction in CC2 is expressed by a chemical equation 3.36,

where fuel for reheater combustion chamber is β (kmol) and ω is the fraction of

steam injection.

(α+1-λ) CO2 + (2α+ζ) H2O + (𝜙-2α) O2 + (𝜃*3.76) N2 + β CH4 + ω H2O

(α+β-λ+1)CO2 + (ζ+ω+2α+2β) H2O + (𝜙-2α-2β) O2 + (𝜃*3.76) N2 (3.36)

Mass of oxidizer = mcc2 = 44*(α+1-λ) CO2 + 18*(2α+ζ) H2O + 32*(𝜙-2α) O2

+ 28*(𝜃*3.76) N2 + 18* ω H2O

(3.37)

Mass of fuel = mf2 = (16𝛽)𝐶𝐻4 (Kg) (3.38)

Oxidizer Fuel ratio = 1.5α + 2𝜙 + 6.58𝜃 + 1.12ζ+1.12ω– 2.75λ + 2.75

β

(3.39)

Gas temperature (Tg2) coming out from CC2 can be found from equation 4.40.

h7 = Cp7 Tg2 (3.40)

41

Total enthalpy of combusted gas (h7) after CC2 can be found from sum of molar

fractions of combustion products and the enthalpy of each gas component in gas

mixture at certain temperature dividing by the molar mass of the exhaust gases.

∑ 𝑦𝑖(ℎ𝑓𝑖

0+ ℎ𝑖

−𝑛

𝑖=0

ℎ0

)𝑟

hr,7 = (3.41) 𝑚𝑟𝑐𝑐2

where, mass of reactant gas of CC2 can be found from eq. 3.36

𝑚𝑟𝑐𝑐2= (α + 1 − λ)CO2 + (2α + ζ)𝐻2𝑂 + (𝜙 − 2𝛼)𝑂2 + (𝜃 ∗ 3.76)𝑁2

+ (𝛽)𝐶𝐻4 + (𝜔)𝑠𝑡𝑒𝑎𝑚 (𝑘𝑚𝑜𝑙)

(3.42)

Mass of product gas of CC2 can be described as,

𝑚𝑝𝑐𝑐2= (α + β − λ + 1)CO2 + [ (2α + 2β + ζ + ω)]H2O

+ (𝜙 − 2α − 2β)O2 + (𝜃 ∗ 3.76)N2 (𝑘𝑚𝑜𝑙)

(3.43)

Eq. 3.22 can be written as eq. 3.44 with taking the reactant gases described in eq.

3.42

ℎ𝑟,7 = {((α + β + 1 − λ)

𝑚𝑟𝑐𝑐2

) (ℎ𝑓0 + ℎ − ℎ0)}

𝐶𝑂27

+ {((2α + 2β + ω + ζ)

𝑚𝑟𝑐𝑐2

) (ℎ𝑓0 + ℎ − ℎ0)}

𝐻2𝑂7

+ {((𝜙 − 2𝛼 − 2𝛽)

𝑚𝑟𝑐𝑐2

) (ℎ𝑓0 + ℎ − ℎ0)}

𝑂27

+ {((𝜃 ∗ 3.76)

𝑚𝑟𝑐𝑐2

) (ℎ𝑓0 + ℎ − ℎ0)}

𝑁27

(3.44)

42

The molar weight of the combustion product (m7) (Kg) is calculated by multiplying

the molars’ fraction by molecular weight for each combustion product component

shows in equation 3.45

m7 = (α+β−λ+1)MCO2 + (2α+2β+ζ+ω)MH2O + (𝜙−2α−2β)MO2 + (𝜃∗3.76)MN2

(α+β−λ+1)CO2 +[ 2(α+β+ζ)]H2O + (𝜙−2α−2β)O2 + (𝜃∗3.76)N2

(3.45)

Specific heat at state 7 (Cp7) (Kg/kmol.K) can be derived similarly as described in

equation 3.28 with the use of equation 3.29 and table 3.2.

Cp7 =

(α+β−λ+1)MCO2CpCO2+ (2α+2β+ζ+ω)MH2OCpH2O

+(𝜙−2α−2β)MO2CpO2+(𝜃∗3.76)MN2CpN2

(α+β−λ+1)CO2 +[ 2(α+β+ζ)]H2O + (𝜙−2α−2β)O2 + (𝜃∗3.76)N2

(3.46)

By trial and error, the Tg2 is set at a specific temperature, which can be achieved by

varying the amount of fuel (β) CH4 in the main combustion chamber and taking in to

account that this temperature should be in a reasonable range between 1200K and

1600K.

To find entropy destruction in CC2, similar procedure can be applied as described

in finding entropy destruction in CC1. For reheater, the equations can be written as,

43

𝑠𝑟𝑒𝑎𝑐𝑡𝑎𝑛𝑡𝑐𝑐2= (α + 1 − λ)sCO2,6 + (2α + ζ)𝑠𝐻2𝑂,6 + (𝜙 − 2α)𝑠𝑂2,6

+ (𝜃 ∗ 3.76)𝑠𝑁2,6 + 𝛽𝑠𝐶𝐻4,1 + 𝜁𝑠𝐻2𝑂,14

− 𝑅𝑢 {(α + 1 − λ)𝑙𝑛 ((α + 1 − λ)P6

𝑚𝑟𝑐𝑐2

)

+ (2α + ζ)𝑙𝑛 ((2α + ζ)P6

𝑚𝑟𝑐𝑐2

) + (𝜙 − 2α)𝑙𝑛 ((𝜙 − 2α)P6

𝑚𝑟𝑐𝑐2

)

+ (𝜃 ∗ 3.76)𝑙𝑛 ((𝜃 ∗ 3.76)P6

𝑚𝑟𝑐𝑐2

) + β𝑙𝑛 (βP6

𝑚𝑟𝑐𝑐2

) + 𝜔𝑙𝑛 (ωP6

𝑚𝑟𝑐𝑐2

)}

(3.47)

𝑠𝑝𝑟𝑜𝑑𝑢𝑐𝑡𝑐𝑐2= (α + β − λ + 1)sCO2,7 + 2(α + β + ζ)𝑠𝐻2𝑂,7

+ (𝜙 − 2α − 2β)𝑠𝑂2,7 + (𝜃 ∗ 3.76)𝑠𝑁2,7

− 𝑅𝑢 {(α + β − λ + 1)𝑙𝑛 ((α + β + λ + 1)P7

𝑚𝑝𝑐𝑐2

)

+ (2α + 2β + ζ + ω)𝑙𝑛 ((2α + 2β + ζ + ω)P7

𝑚𝑝𝑐𝑐2

)

+ (𝜙 − 2α − 2β)𝑙𝑛 ((𝜙 − 2α − 2β)P7

𝑚𝑝𝑐𝑐2

)

+ (𝜃 ∗ 3.76)𝑙𝑛 ((𝜃 ∗ 3.76)P7

𝑚𝑝𝑐𝑐2

)}

(3.48)

𝑠𝑔𝑒𝑛𝑐𝑐2= 𝑠𝑝𝑟𝑜𝑑𝑢𝑐𝑡𝑐𝑐2

− 𝑠𝑟𝑒𝑎𝑐𝑡𝑎𝑛𝑡𝑐𝑐2

(3.49)

The exergy destruction in CC2 can be found from equation 3.50.

��d𝑐𝑐2= ��7𝑇0 𝑠𝑔𝑒𝑛𝑐𝑐2

(3.50)

Exegy efficiency of CC2,

44

𝜂𝑒𝑥𝐶𝐶2= 1 −

𝐼��𝑐𝑐2

��7

(3.51)

where,

��7 = ��7[(ℎ7 − ℎ0) − 𝑇0(𝑠𝑝𝑟𝑜𝑑𝑢𝑐𝑡𝑐𝑐2−𝑠0) (3.52)

3.3.3 Natural gas fired gas turbines (GT1 and GT2)

The pressure ratio (Prg) represents the pressure difference between state 8 and

state 5.

Prg =𝑃8

𝑃5

(3.53)

The optimum intermediate pressure for expansion is calculated as given below.

P6 = √P5 x P8 (3.54)

Since the relative pressure (P6) has been determined, the absolute entropy (s6s),

ideal enthalpy (h6s) be obtain from EES. The actual enthalpy at GT1 outlet (state 6)

is calculated considering gas turbine isentropic efficiency (η𝑔).

h6 = h5 – η𝑔 (h5 – h6s) (3.55)

Molar fractions of combustion products at state (5) and state (6) remains the same,

trial and error method is used to estimate GT1 outlet temperature (T6) based on the

value of h6 found from equation 3.55. Similarly enthalpy at state (8) can be found as

described in equation 3.56.

45

h8 = h7 – η𝑔 (h7 – h8s) (3.56)

By using energy balance equations for GT1 and GT2, we find and work output

obtained after gas expansion at desired pressure ratio.

��𝐺𝑇1 = ��5(ℎ5 − ℎ6) (3.57)

��𝐺𝑇2 = ��7(ℎ7 − ℎ8)

(3.58)

Physical exergy rate of GT1 and GT2,

��5 = ��5[(ℎ5 − ℎ0) − 𝑇0(𝑠5−𝑠0)] (3.59)

��6 = ��5[(ℎ6 − ℎ0) − 𝑇0(𝑠6−𝑠0)] (3.60)

��7 = ��7[(ℎ7 − ℎ0) − 𝑇0(𝑠7−𝑠0)]

(3.61)

��8 = ��7[(ℎ8 − ℎ0) − 𝑇0(𝑠8−𝑠0)]

(3.62)

Exergy destruction of GT1 and GT2 can be found from the exergy balance equations

described as,

𝐼��𝑇1 = ��5 − ��6 − ��𝐺𝑇1 (3.63)

𝐼��𝑇2 = ��7 − ��8 − ��𝐺𝑇2 (3.64)

Exergy efficiency of GT1 and GT2,

46

𝜂𝑒𝑥𝐺𝑇1=

��𝐺𝑇1

��5− ��6

(3.65)

𝜂𝑒𝑥𝐺𝑇2=

��𝐺𝑇2

X7−X8

(3.66)

3.3.4 Heat recovery steam generator (HRSG)

The gas turbine cycle couples with the steam cycle through a heat exchanger which

knows as a heat recovery steam generator (HRSG). In its simplest form, is

continuous tube heat exchanger in which preheating, evaporation and superheating

of the feed water take place consecutively as shown in figure 3.2. Many tubes are

mounted in parallel and are joined by headers thus providing a common inlet for

feed water and a common outlet for steam. Water is forced at the cold end of HRSG

(state 13). The water changes phase along the circuit and exists as superheated

steam at state 10. The exhaust gas from CC2 passes through HRSG from state 8

to state 9, opposite direction to that of the water flow. It is assumed that heat

recovery steam generator (HRSG) is working adiabatically.

Figure 3.3 Configuration of Heat Recovery Steam Generator (HRSG)

Economizer Superheater Evaporator ‘13’

Feed water

‘10’ Steam

‘9’ Gas out

‘8’ Hot Gas

47

Since enthalpy at state 8 (h8) is knows by equation 3.56 and temperature at state 8

(T8) can be found from trial and error method described similarly in the case of

combustion chambers. Amount of steam generated from HRSG can be found from

energy balance equation.

𝑚8(ℎ8 − ℎ9) = 𝑚10(ℎ10 − ℎ13) (3.67)

Energy efficiency of HRSG can be described as,

𝜂𝐻𝑅𝑆𝐺 = (ℎ10 − ℎ13)

(ℎ8 − ℎ9)

(3.68)

Exergy destruction of HRSG,

𝐼��𝑅𝑆𝐺 = (��8 − ��9) − (��10 − ��13) (3.69)

Exergy efficiency of HRSG,

𝜂𝑒𝑥𝐻𝑅𝑆𝐺=

(��10 − ��13)

(��8 − ��9)

(3.70)

3.3.4 Steam turbine (ST)

Steam generated from HRSG expanded through ST. A mass fraction of steam 𝜻 &

ω are extracted from ST at a pressure 5% higher than the pressure in CC1 and CC2

respectively. There are total of two streams of steam fraction, that means total mass

of (𝜻 + 𝝎) is less available for the ST to carry out work. Extracted steam is injected

48

to CC1 and CC2 equally at a mass fraction of 𝜻 & ω respectively. Steam turbine

work can be obtained from energy balance equation.

��𝑠𝑡 = ��10(ℎ10 − ℎ11) − ��14(ℎ14 − ℎ11) (3.71)

Where,

��11 = ��10 − ��14 (3.72)

��14 = (𝜁 + 𝜔)𝑠𝑡𝑒𝑎𝑚 ∗ 18 (𝐾𝑔/𝑠) (3.73)

Exergy at each stage,

��10 = ��10[(ℎ10 − ℎ0) − 𝑇0(𝑠10 − 𝑠0)] (3.74)

��11 = ��11[(ℎ11 − ℎ0) − 𝑇0(𝑠11 − 𝑠0)] (3.75)

��14 = (𝜁 + 𝜔)𝑠𝑡𝑒𝑎𝑚 ∗ 18[(ℎ14 − ℎ0) − 𝑇0(𝑠14 − 𝑠0)] (3.76)

Exergy destruction in ST,

𝐼��𝑇 = ��10 − (��11 + ��14 + ��𝑆𝑇) (3.77)

Exergy efficiency of ST,

𝜂𝑒𝑥𝑆𝑇=

��𝑆𝑇

(��10 − ��11 − ��14)

(3.78)

3.3.5 Condenser (CON)

Exergy destruction in CON,

49

𝐼��𝑂𝑁 = ��11 − ��12 (3.79)

3.3.6 Water circulation pump for bottoming cycle

Saturated water coming out of condenser is pressurised by a pump. Pressure at

state 13 and make up water supplied at condensate temperature are assumed.

Actual enthalpy at state 13 can be found pump isentropic efficiency.

h13 = h13s− h12

η𝑐 + h12

(3.80)

Pump work can be found from energy balance across pump,

��𝑃 = ��𝑠(ℎ13 − ℎ12) (3.81)

Exergy destruction of pump,

𝐼�� = ��12 + ��𝑃 − ��13 (3.82)

Exergy efficiency of pump,

𝜂𝑒𝑥𝑃= 1 −

𝐼��

��13

(3.83)

Overall natural gas combine cycle efficiencies

For configuration 1, 2 and 3:

Total work output,

50

��𝑁𝐸𝑇1 = ��𝐺𝑇1 + ��𝐺𝑇2 + ��𝑆𝑇 − (��𝐶1 + ��𝐶2 + ��𝑃) (3.84)

Combined cycle energy efficiency,

𝜂𝑎𝑖𝑟 =��𝑁𝐸𝑇1

��𝑓𝑢𝑒𝑙 ∗ 𝐿𝐻𝑉

(3.85)

Combine cycle exergy efficiency,

𝜂𝑒𝑥𝑎𝑖𝑟=

��𝑁𝐸𝑇1

��𝑓𝑢𝑒𝑙

(3.86)

For oxyfuel combined cycle configuration 4, 5 & 6,

Total work output from oxyfuel combine cycle can be carried out with equation 3.87.

��𝑁𝐸𝑇2 = ��𝐺𝑇1 + ��𝐺𝑇2 + ��𝑆𝑇 − (��𝐶1 + ��𝐶2 + ��𝑃 + (𝜆 ∗ 0.092) + (𝜙

∗ 0.225)

(3.87)

Oxyfuel combined cycle power generation thermal efficiency,

𝜂𝑜𝑥𝑦 =��𝑁𝐸𝑇2

��𝑓𝑢𝑒𝑙 ∗ 𝐿𝐻𝑉

(3.88)

Oxyfuel combined cycle power generation exergy efficiency,

𝜂𝑒𝑥𝑜𝑥𝑦=

��𝑁𝐸𝑇2

��𝑓𝑢𝑒𝑙

(3.89)

51

Chapter 4: Results and discussions

The performance analysis of combine cycles power plant has been studied and

presented here. There are two basic cycles, one where fuel is oxidised with air and

the other one being oxidized by oxygen only. All simulation were run on computer

program “Engineering Equation Solver (EES)”. The program code has been

validated by comparing the results of published literatures. A hand calculation is

also carried out at base condition and compared with the EES result to validate the

program.

4.1 Component range of values used in studies

The pressure drop for the air intercooler (IC), combustion chambers (CC1

& CC2) and heat recovery steam generator (HRSG) is 2%

The compressor isentropic efficiency η𝑐 is 87%[8]

The gas turbine isentropic efficiency η𝐺𝑇 is 89%[8]

The steam pressure at the steam turbine inlet is 100 bar[12]

The steam temperature at the steam turbine inlet is 500° C[12]

Steam turbine exhaust pressure is 0.1 bar [14]

The stack temperature is 85° C [14]

The pump isentropic efficiency is 0.85% [8]

The steam turbine isentropic efficiency is 0.85% [8]

The fuel is methane gas, which has a lower heating value of 42,000 kJ/kg

The ambient pressure (P0) and temperature (T0) are, respectively, 1 bar and

25° C

Air is composed of 21% O2 and 79% N2 [23]

Energy required to compress and liquefy CO2 is 0.092 kWh/kg [26]

52

Energy required to separate O2 from air through ASU at ambient condition

is 225 kWh/tone [27]

Purity of O2 separated from ASU is considered 100%

Natural gas consists 100% methane

Minimum temperature difference between the flue gas and the steam in

HRSG is taken as 25° C

4.2 Effect of pressure ratio on combined cycle performance with fraction of

steam injection

Table 4.1 Operating parameters of combined cycle configuration for various

pressure ratio (configuration 1, 2 and 3)

TIT1 & TIT2 Theoretical

Air (%)

Ambient

Temp

Ambient

Pressure

Steam

Temperature

Steam

Pressure

1200° C 200 25° C 1 Bar 500° C 100 Bar

40

42

44

46

48

50

52

54

56

58

60

5 1 0 1 5 2 0 2 5 3 0 3 5 4 0

THER

MA

L EF

FIC

IEN

CY

(%)

PRESSURE RATIO

0% Steam Injection 5% Steam Injection in CC1 5% Steam Injection in CC1 & CC2

53

Figure 4.1 Effect of pressure ratio on combined cycle thermal efficiency

When combined cycle operating conditions are set as described in table 4.1, the

thermal efficiency of the combined cycles shows in figure 4.1 when pressure ratio

range is selected from 5 to 40. There is a sharp increases in thermal efficiency for

initial pressure ratio change from 5 to 25. For configuration 1, thermal efficiency

slightly decreases from pressure ratio of 25 to 40 where in configuration 2 and 3,

there is no decrease in thermal efficiency after optimum pressure ratio reached.

Figure 4.2 shows effect of pressure ratio on exergy efficiency. Overall there is an

increase of 1.57 % exergy efficiency with 5 % steam injection in CC1 & CC2.

Figure 4.2 Effect of pressure ratio on combined cycle exergy efficiency

Effect of pressure ratio on the topping cycle and steam cycle work output is shown

in figure 4.3 and 4.4 respectively. Pressure ratio varied from 5 to 40 and gas turbine

inlet temperature for both turbines are fixed at 1200° C. Topping cycle shows 17%

improvement on power output with 5 % steam injection in CC1 at lower pressure

ratio of 5. When 5 % steam is injected in CC2 as well, there is a further improvement

of 8 % in topping power output. Increase in power output stabilizes after pressure

40

42

44

46

48

50

52

54

56

58

5 1 0 1 5 2 0 2 5 3 0 3 5 4 0

EXER

GY

EFFI

CIE

NC

Y

PRESSURE RATIO

0% Steam Injection 5% Steam Injection in CC1 5% Steam Injection in CC1 & CC2

54

ratio of 10. The pressure ratio shows opposite effect on the steam cycle work output

with the gas turbine inlet temperature fixed, increasing the pressure ratio led to lower

exhaust gas temperature after second gas turbine. This led to reduced steam

generation in the heat recovery steam generator, hence, lower work output is

produce. Steam turbine work further lowers when ζ is injected in CC1 and ω is

injected in CC2. However figure 4.5 shows increase in combined cycle power output

with steam injection in CC1 & CC2 together as increase in topping cycle outcomes

the decrease in steam cycle.

Figure 4.3 Effect of pressure ratio on the topping cycle work output with 5%

steam injection

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400000

5 1 0 1 5 2 0 2 5 3 0 3 5 4 0

TOP

PIN

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WO

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OU

TPU

T (K

W)

PRESSURE RATIO

0% Steam Injection 5% Steam Injection in CC1 5% Steam Injection in CC1 & CC2

55

Figure 4.4 Effect of pressure ratio on the steam cycle work output with 5 %

steam injection

Figure 4.5 Effect of pressure ratio on the combined cycle work output with 5

% steam injection

100000

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200000

5 1 0 1 5 2 0 2 5 3 0 3 5 4 0

STEA

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T (K

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PRESSURE RATIO

0% Steam Injection 5% Steam Injection in CC1 5% Steam Injection in CC1 & CC2

300000

320000

340000

360000

380000

400000

420000

440000

460000

480000

5 1 0 1 5 2 0 2 5 3 0 3 5 4 0

CO

MB

INED

CYC

LE W

OR

K O

UTP

UT

(KW

)

PRESSURE RATIO

0% Steam Injection 5% Steam Injection in CC1 5% Steam Injection in CC1 & CC2

56

4.3 Effect on CO2 emission with fraction of steam injection

Table 4.2 Operating parameters of combined cycle configuration to

determine CO2 emission (configuration 1, 2 and 3)

Pressure

Ratio

Ambient

Temp

Ambient

Pressure

Steam

Temperature

Steam

Pressure

25 25° C 1 Bar 500° C 100 Bar

Amount of fuel reduction with steam injection as direct impact on carbon dioxide

emission. For example, 1 kg of methane reduces 2.75 kg of carbon dioxide. When

operating parameters of the combined system are set as described in table 4.2, the

effect of steam injection on CO2 emission is shown in figure 4.6. When ζ is injected

to CC1, a constant TIT1 is maintained through reduction on air mass flow thus

reduction in compressor work. At this air flow rate, combined cycle net work output

increase by 3.2 % due to increase in mass flow of the steam from 0% to 10% in

GT1. To maintain constant combined cycle work output, there is a reduction in fuel

consumption alpha. Overall 3.4 % CO2 emission reduced at 10% steam injection.

When ω injected in CC2 only, mass flow rate affects only GT2 and through fuel β

reduction in CC2, 7.2 % of CO2 emission reduction observed when steam injected

from 0% to 10%. However when ζ and ω injected in CC1 and CC2 together, only

0.9 % of CO2 emission reduction observed. Injecting steam in CC2 only has greater

effect on reduction of carbon dioxide emission.

57

Figure 4.6 Effect of steam injection in CC1 and CC2 on Carbon Dioxide

Emission

4.4 Effect of ambient temperature on combined cycle performance with

fraction of steam injection

Table 4.3 Operating parameters of combined cycle configuration for variable

ambient temperature (configuration 1, 2 and 3)

Pressure

Ratio

TIT1 &

TIT2

Theoretical

Air (%)

Ambient

Pressure

Steam

Temperature

Steam

Pressure

25 1200° C 200 1 Bar 500° C 100 Bar

0.104

0.106

0.108

0.11

0.112

0.114

0.116

0.118

1 2 3 4 5 6 7 8 9 1 0

CA

RB

ON

DIO

XID

E EM

ISSI

ON

(KG

/MW

)

FRACTION OF STEAM INJECTION (%)

ζ

ζ & ω

ω

58

When operating parameters of the combined cycle system set as described in table

4.3, effect on work output studies with ambient temperature of 5° C, 25° C and 50°

C. Figure 4.7 shows the work output in topping cycle and steam cycle at various

ambient temperature for configuration 2 and figure 4.8 shows work output for

configuration 3. Ambient air has significant influence on the power output and

efficiency of gas turbine. Increasing the ambient temperature reduces the density of

the air and consequently reduces the air mass flow into the compressor as constant

volume engine. The specific power consumed by compressor increases

proportionally to the air intake temperature without a corresponding increase in the

output from the turbine part. However combine cycle work increases as higher

temperature at gas turbine exhaust increases work output of the steam cycle.

Figure 4.7 Effect of ambient air temperature on cycle work output with steam

for configuration 2

80000

130000

180000

230000

280000

330000

1 2 3 4 5 6 7 8 9 1 0

WO

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OU

TPU

T (K

W)

FRACTION OF STEAM INJECTION IN CC1 (%)

Topping cycle work @ T1 = 50° C Topping cycle work @ T1 = 25° CTopping cycle work @ T1 = 5° C Steam cycle work output @ T1 = 5° CSteam cycle work output @ T1 = 25° C Steam cycle work output @ T1 = 50° C

59

Figure 4.8 Effect of ambient air temperature on cycle work output for

configuration 3

4.5 Performance analysis on fuel ratio on combined cycle with fraction of

steam injection

Table 4.4 Fuel balance on CC1 and CC2 for configuration 1

Theoair (%)

α (%) ζ

(%) TIT (°C)

β (%) ω

(%) TIT2 (°C)

ηTH (%)

WC

(kW) WST

(kW) WNET (kW)

200 40 0 1053 60 0 1222 43.38 250363 146469 348051

200 42.22 0 1074 57.78 0 1217 44.01 250363 145701 353104

200 44.44 0 1095 55.56 0 1213 44.64 250363 144921 358139

200 46.67 0 1116 53.33 0 1208 45.26 250363 144128 363155

200 48.89 0 1137 51.11 0 1204 45.89 250363 143321 368153

200 51.11 0 1157 48.89 0 1199 46.51 250363 142502 373130

200 53.33 0 1178 46.67 0 1194 47.13 250363 141670 378088

200 55.56 0 1198 44.44 0 1189 47.74 250363 140824 383024

200 57.78 0 1218 42.22 0 1184 48.35 250363 139966 387939

200 60 0 1238 40 0 1179 48.96 250363 139094 392832

80000

130000

180000

230000

280000

330000

1 2 3 4 5 6 7 8 9 1 0

WO

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TPU

T (K

W)

FRACTION OF STEAM INJECTION IN CC1 & CC2 (%)

Topping cycle work @ T1 = 50° C Topping cycle work @ T1 = 25° C

Topping cycle work @ T1 = 5° C Steam cycle work output @ T1 = 5° C

Steam cycle work output @ T1 = 25° C Steam cycle work output @ T1 = 50° C

60

Table 4.4 shows fuel supply ratio between CC1 & CC2. At Pr of 25 & Theoair of 200

%, ideal fuel supply found is α = 54% & β = 46%. At this fuel supply, TIT1 & TIT2

are very close to each other which is an important factor to gas turbine efficiency.

Table 4.5 shows steam ζ & ω injected at CC1 & CC2 respectively. As stem injected

(ζ) in CC1, there is a reduction in TIT1 due to increased mass from steam resulting

flame temperature down. There is a sharp reduction TIT of 88°C with 10% steam

injection in CC1. Further steam addition ω in CC2 brings TIT2 even further down.

Work output of the gas turbines are increased about 5 MW but sharp decline in net

work out put (37 MW) due to work output lost from steam turbine. Table 4.5 shows

ideal theoretical air input at different steam injection to maintain the same net work

output. Theoretical air reduced by 32 %. It is assumed that complete combustion

takes place in the CC1 & CC2 due to excess air in the combustion chamber. Steam

injection decreases the amount of excess air in the combustion chamber as it also

controls the temperature. Air should not decrease below the minimum requirement

for the complete combustion of the fuel. At proper balance of fuel supply α (54%) &

β (46%), theoretical air must not go below required (𝜙-2α-2β). It implies that the

amount of steam injection has a limit depending on the air quantity in the

compressor. Steam mass ratio corresponding to this value is 5%. As a result of

decreased fuel mass, the flue gas from the combustion chamber decreases resulting

lower CO2 & NO2 emission.

61

Table 4.5 Work output of combined cycle for configuration 3

Theoair (%)

ζ (%)

TIT (°C)

ω

(%)

TIT2 (°C)

ηTH (%)

WC

(kW) WGT

(kW) WST

(kW) WNET (kW)

200 1 1173 1 1169 46.93 250363 491503 136722 376533

200 2 1162 2 1146 46.52 250363 492837 132014 373246

200 3 1152 3 1123 46.08 250363 493950 127296 369729

200 4 1142 4 1102 45.62 250363 494858 122566 365996

200 5 1132 5 1081 45.13 250363 495575 117824 362060

200 6 1122 6 1061 44.61 250363 496115 113068 357934

200 7 1112 7 1042 44.08 250363 496490 108297 353627

200 8 1103 8 1023 43.52 250363 496710 103510 349152

200 9 1094 9 1005 42.94 250363 496787 98707 344516

200 10 1085 10 987.5 42.34 250363 496729 93886 339730

Figure 4.9 shows the topping cycle, steam cycle and combined cycle work output

when steam is injected in CC1 only (configuration 2), steam is injected in both CC1

& CC2 together (configuration 3)and steam injection in CC2 only. At fixed pressure

ratio of 25 and TIT1 & TIT2 fixed at 1000° C, In the case of steam injection in CC1

only, combined power output increased 18 % from no steam injection to 10% steam

injection. When steam injected in CC2 only with the same operating condition,

however work output increased to 9.5% only. It is more beneficial to add steam in

CC1 instead of CC2 when there is a choice to inject steam in only one combustion

chamber.

62

Figure 4.9 Effect on work output with steam injection for configuration 1, 2

and 3

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330000

1 2 3 4 5 6 7 8 9 10

WO

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TPU

T (K

W)

FRACTION OF STEAM INJECTION (%)

Gas Cycle Work output (Steam Injection in CC1)

Steam Cycle work output (Steam Injection in CC1)

Combine cycle work output (Steam Injection in CC1)

Gas Cycle Work output (Steam Injection in CC1 & CC2)

Steam Cycle work output (Steam Injection in CC1 & CC2)

Combine cycle work output (Steam Injection in CC1 & CC2)

Gas Cycle Work output (Steam Injection in CC2)

Steam Cycle work output (Steam Injection in CC2)

Combine cycle work output (Steam Injection in CC2)

63

4.6 Effect of TIT on performance of combined cycle with fraction of steam

injection

Table 4.6 Operating parameters of combined cycle configuration for various

turbine inlet temperature (configuration 1, 2 & 3)

Pressure

Ratio

Air Mass

flow rate

Theoretical

Air (%)

Ambient

Temp

Ambient

Pressure

Steam

Temperature

Steam

Pressure

25 300

(Kg/s)

200 25° C 1 Bar 500° C 100 Bar

When operating conditions of the combined cycle system is set as described in table

4.6, work output of topping, steam and combined cycle studied with turbine inlet

temperature range of 1000° C to 1400° C. This result is obtained with TIT being fixed

at particular temperature by adding more fuel in the system that means when steam

is injected in the combustion chamber, if no fuel is added, TIT would lower because

of the additional steam mass reacting with products of combustion chamber at

relative low temperature, bringing entire combustion chamber temperature down.

Figure 4.10 shows the work output of topping, steam and combined cycle for

configuration 1. For a particular pressure ratio, an increase in gas turbine inlet

temperature led to increase in combined cycle net work output. Figure 4.11 shows

effect of TIT on topping cycle work output with various increments of steam injection

for configuration 2. Higher steam injection and higher TIT leads to maximum work

output on the topping cycle. Figure 4.12 shows 11% increase in topping cycle for

configuration 3.

64

Figure 4.10 Effect of various TIT on combined cycle work output for configuration 1

Figure 4.11 Effect of TIT on topping cycle work output for configuration 2

0

100000

200000

300000

400000

500000

600000

700000

1 0 0 0 1 1 0 0 1 2 0 0 1 3 0 0 1 4 0 0

WO

RK

OU

TPU

T (K

W)

TIT1 & TIT2 ( C )

Combined cycle work output

Gas cycle work output

Steam cycle work output

100000

150000

200000

250000

300000

350000

400000

450000

500000

550000

1 2 3 4 5 6 7 8 9 1 0

TOP

PIN

G C

YCLE

WO

RK

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TPU

T (K

W)

FRACTION OF STEAM INJECTION IN CC1 (%)

TIT1 & TIT2 = 1400° C

TIT1 & TIT2 = 1300° C

TIT1 & TIT2 = 1200° C

TIT1 & TIT2 = 1100° C

TIT1 & TIT2 = 1000° C

65

Figure 4.12 Effect of TIT on topping cycle work output for configuration 3

Higher TIT leads to higher temperature of the exhaust gas that enters heat recovery

steam generator, thus producing more steam for the steam cycle. Figure 4.13 & 4.14

shows work output of steam cycle for configuration 2 and 3 respectively. At lower

TIT of 1000° C, steam injection for both configuration 2 and configuration 3 shows

decline in workout put as more steam is injected compared to the higher TIT of 1400°

C because lower TIT yields to lower entry temperature at heat recovery steam

generator and producing lower amount of steam. Efficiency of heat recovery steam

generator is lower at lower temperature. As more steam is injected in topping cycle,

it further lowers the exhaust gas temperature that leads to overall lower work output

of steam cycle. At 1400° C, the decline in steam cycle work output with higher

percentage of steam injection has lower impact. Figure 4.15 and 4.16 shows

combined cycle work output for configuration 2 and 3 respectively. Despite of

reduction in steam cycle with more steam injection, combined cycle work output

shows improvement for configuration 3 as topping cycle produces more power.

100000

150000

200000

250000

300000

350000

400000

450000

500000

550000

600000

1 2 3 4 5 6 7 8 9 1 0

TOP

PIN

G C

YCLE

WO

RK

OU

TPU

T (K

W)

FRACTION OF STEAM INJECTION IN CC1 & CC2 (%)

TIT1 & TIT2 = 1000° C

TIT1 & TIT2 = 1100° C

TIT1 & TIT2 = 1200° C

TIT1 & TIT2 = 1300° C

TIT1 & TIT2 = 1400° C

66

Figure 4.13 Effect of TIT on steam cycle work output for configuration 2

Figure 4.14 Effect of TIT on steam cycle work output for configuration 3

60000

80000

100000

120000

140000

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1 2 3 4 5 6 7 8 9 1 0

STEA

M C

YCLE

WO

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TPU

T (K

W)

FRACTION OF STEAM INJECTION IN CC1 (%)

TIT1 & TIT2 = 1000° C

TIT1 & TIT2 = 1100° C

TIT1 & TIT2 = 1200° C

TIT1 & TIT2 = 1300° C

TIT1 & TIT2 = 1400° C

60000

80000

100000

120000

140000

160000

180000

200000

1 2 3 4 5 6 7 8 9 1 0

STEA

M C

YCLE

WO

RK

OU

TPU

T (K

W)

FRACTION OF STEAM INJECTION IN CC1 & CC2 (%)

TIT1 & TIT2 = 1000° C

TIT1 & TIT2 = 1100° C

TIT1 & TIT2 = 1200° C

TIT1 & TIT2 = 1300° C

TIT1 & TIT2 = 1400° C

67

Figure 4.15 Effect of TIT on combined cycle work output with fraction of

steam injection in CC1

Figure 4.16 Effect of TIT on combined cycle work output with fraction of

steam injection in CC1 and CC2

100000

200000

300000

400000

500000

600000

700000

800000

1 2 3 4 5 6 7 8 9 1 0

CO

MB

INED

CYC

LE W

OR

K O

UTP

UT

(KW

)

FRACTION OF STEAM INJECTION IN CC1 (%)

TIT1 & TIT2 = 1000° C

TIT1 & TIT2 = 1100° C

TIT1 & TIT2 = 1200° C

TIT1 & TIT2 = 1300° C

TIT1 & TIT2 = 1400° C

100000

200000

300000

400000

500000

600000

700000

800000

1 2 3 4 5 6 7 8 9 1 0

CO

MB

INED

CYC

LE W

OR

K O

UTP

UT

(KW

)

FRACTION OF STEAM INJECTION IN CC1 & CC2 (%)

TIT1 & TIT2 = 1000° C

TIT1 & TIT2 = 1100° C

TIT1 & TIT2 = 1200° C

TIT1 & TIT2 = 1300° C

TIT1 & TIT2 = 1400° C

68

4.7 Effect of Isentropic efficiencies on performance of combined cycle with

fraction of steam injection

Table 4.7 Operating parameters of combined cycle configuration for variable

isentropic efficiency (configuration 1, 2 and 3)

Pressure

Ratio

Air Mass

flow rate

Theoretical

Air (%)

Ambient

Temp

Ambient

Pressure

Steam

Temperature

Steam

Pressure

25 300

(Kg/s)

200 25° C 1 Bar 500° C 100 Bar

Figure 4.17 Effect of Isentropic efficiency on Topping cycle work output with

fraction of steam injection in CC1

200000

210000

220000

230000

240000

250000

260000

270000

280000

290000

300000

1 2 3 4 5 6 7 8 9 1 0

TOP

PIN

G C

YCLE

WO

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OU

TPU

T (K

W)

FRACTION OF STEAM INJECTION IN CC1

Isentropic Efficincy = 85%

Isentropic Efficincy = 90%

Isentropic Efficincy = 95%

69

Figure 4.18 Effect of Isentropic efficiency on bottom cycle work output with

fraction of steam injection in CC1

Figure 4.19 Effect of Isentropic efficiency on combined cycle work output

with fraction of steam injection in CC1

100000

110000

120000

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1 2 3 4 5 6 7 8 9 1 0

BO

TTO

M C

YCLE

WO

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T (K

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FRACTION OF STEAM INJECTION IN CC1

Isentropic Efficincy = 85%

Isentropic Efficincy = 90%

Isentropic Efficincy = 95%

330000

350000

370000

390000

410000

430000

450000

1 2 3 4 5 6 7 8 9 1 0

CO

MB

INED

CYC

LE W

OR

K O

UTP

UT

KW

)

FRACTION OF STEAM INJECTION IN CC1

Isentropic Efficincy = 85%

Isentropic Efficincy = 90%

Isentropic Efficincy = 95%

70

4.8 Effect of flue gas recycle on turbine inlet temperatures for configuration

4, 5 and 6 (oxyfuel combustion)

When fuel is burned with pure oxygen, it produces very high TIT which is not

desirable for current operational turbine blades. To bring the TIT at operable level,

the flue gas (λ) is recycled to the combustion chamber once all water vapour is

isolated. Table 4.3 shows the effect of (λ) on TIT1 & TIT2 when fuel supply is

maintained at α = 54% & β = 46%

Table 4.8 Effect of flue gas recycle on turbine inlet temperature in combined cycle system for configuration 4

λ (%)

TIT1 (° C)

TIT2 (° C)

90 1552 1619

80 1643 1697

70 1748 1784

60 1871 1880

50 2017 1987

40 2194 2105

30 2412 2237

20 2687 2383

10 3042 2540

0 3545 2710

4.9 Effect of steam injection on TIT with fraction of flue gas recycle for

configuration 4 (Oxyfuel combustion)

Figure 4.20 shows the effect of steam injection (5% & 10% respectively) only in

CC1. Steam addition helps to lower the TIT as it adds mass of steam in

combustion chamber. Figure 4.21 shows the effect of steam injection (5% & 10%

71

respectively) in CC1 & CC2 together. As more steam is injected in CC1, additional

mass flow from steam reduces the TIT and less recycled flue gas required to bring

the TIT to the operational level of 1000° C – 1400° C. However, figure 4.22 shows

combined cycle efficiency is maximum when steam injected only in CC1 and

minimum when steam is injected in CC1 & CC2 together. In configuration 2 where

5% steam injected only in CC1, optimum lambda is 70% when thermal efficiency is

the highest (73.4%) and when 5% steam is injected in CC1 & CC2 together,

maximum thermal efficiency is 73.3% when lambda is 60%. Efficiencies are much

higher compared to results obtained from configurations 2 & 3 because energy

used for air separation unit has not been considered.

Figure 4.20 Effect of steam injection on turbine inlet temperature with fraction of flue gas recycle for configuration 5

0

500

1000

1500

2000

2500

3000

3500

4000

9 0 8 0 7 0 6 0 5 0 4 0 3 0 2 0 1 0 0

TIT

(C)

FRACTION OF FLUE GAS RECYCLE (%)

TIT1 @ ζ = 5%, ω = 0%

TIT2 @ ζ = 5%, ω = 0%

TIT1 @ ζ = 10%, ω = 0%

TIT2 @ ζ =10%, ω = 0%

72

Figure 4.21 Effect of steam injection on turbine inlet temperature with

fraction of flue gas recycle for configuration 6

Figure 4.22 Effect of steam injection on combined cycle thermal efficiency with fraction of flue gas recycle for configuration 5 & 6

0

500

1000

1500

2000

2500

3000

3500

4000

9 0 8 0 7 0 6 0 5 0 4 0 3 0 2 0 1 0 0

TIT

(C)

FRACTION OF FLUE GAS RECYCLE (%)

TIT1 @ ζ = 5%, ω = 5%

TIT2 @ ζ = 5%, ω = 5%

TIT1 @ ζ = 10%, ω = 10%

TIT2 @ ζ = 10%, ω = 10%

62

64

66

68

70

72

74

76

9 0 8 0 7 0 6 0 5 0 4 0 3 0 2 0 1 0 0

THER

MA

L EF

FIC

IEN

CY

(%)

FRACTION OF FLUE GAS RECYCLE (%)

Thermal efficiency @ ζ = 5%, ω = 0%

Thermal efficiency @ ζ = 10%, ω = 0%

Thermal efficiency @ ζ = 10%, ω = 10%

Thermal efficiency @ ζ = 5%, ω = 5%

73

4.10 Effect of pressure ratio on combined cycle work output, efficiencies &

CO2 emission for configurations 4, 5 and 6 (Oxyfuel combustion)

Various pressure ratio applied to the oxyfuel combustion system at fixed TIT1 and

TIT2 of 1200° C. Figure 4.23 shows topping cycle work output. Topping cycle

increases work output up to pressure ratio of 25 then remains constant at higher

pressure ratio. Adding 5% steam injection at CC1 further increases the topping work

output and 5% steam injection at CC1 and CC2 brings topping cycle work slightly

higher. There is an increase of 3% work output when steam injected in CC1 only

and 2% further increase when steam is injected in both combustion chambers.

However, figure 4.24 shows decline in work output when pressure ratio increases.

At higher pressure ratio, there is more excessive oxygen which brings the flue gas

temperature down and less steam is produced. The overall efficiency have been

compared with different thermodynamic parameters. Figure 4.26 and 4.27 shows

thermal and exergy efficiency respectively. Thermal efficiency stabilize at pressure

ratio of 20 while exergy efficiency stabilize at pressure ratio of 18. Figure 4.29 shows

CO2 emission for configuration 4, 5 and 6. The lowest emission found for

configuration 6 when pressure ratio is 25 and steam injection is 5%. Emission at that

condition is 0.08587 Kg/MW. CO2 emission for configuration 1, 2 and 3 at the same

parameter shown in figure 4.6 is 0.1126 Kg/MW which is much lower than result

obtained from configuration 4, 5 and 6 shown in figure 4.28.

74

Figure 4.23 Effect of pressure ratio on topping cycle work output for

configurations 4, 5 and 6

Figure 4.24 Effect of pressure ratio on steam cycle work output for

configurations 4, 5 and 6

190000

200000

210000

220000

230000

240000

250000

5 1 0 1 5 2 0 2 5 3 0 3 5 4 0

TOP

PIN

G C

YCLE

WO

RK

OU

TPU

T (K

W)

PRESSURE RATIO

ζ = 0%

ζ = 5%

ζ = 5%, ω=5%

20000

25000

30000

35000

40000

45000

50000

55000

5 1 0 1 5 2 0 2 5 3 0 3 5 4 0

STEA

M C

YCLE

WO

RK

OU

TPU

T (K

W)

PRESSURE RATIO

ζ = 0%

ζ = 5%

ζ = 5%, ω=5%

75

Figure 4.25 Effect of pressure ratio on combined cycle work output for

configurations 4, 5 and 6

Figure 4.26 Effect of pressure ratio on combined cycle thermal efficiency for

configurations 4, 5 and 6

240000

245000

250000

255000

260000

265000

270000

275000

280000

5 1 0 1 5 2 0 2 5 3 0 3 5 4 0

CO

MB

INED

CYC

LE W

OR

K O

UTP

UT

(KW

)

PRESSURE RATIO

ζ = 0%

ζ = 5%

ζ = 5%, ω=5%

57

58

59

60

61

62

63

64

65

5 1 0 1 5 2 0 2 5 3 0 3 5 4 0

THER

MA

L EF

FIC

IEN

CY

(%)

PRESSURE RATIO

ζ = 0%

ζ = 5%

ζ = 5%, ω=5%

76

Figure 4.27 Effect of pressure ratio on exergy efficiency for configurations 4, 5 and 6

Figure 4.28 Effect of pressure ratio on CO2 emission for configurations 4, 5 and 6

54

55

56

57

58

59

60

61

62

5 1 0 1 5 2 0 2 5 3 0 3 5 4 0

EXER

GY

EFFI

CIE

NC

Y (%

)

PRESSURE RATIO

ζ = 0%

ζ = 5%

ζ = 5%, ω=5%

0.082

0.084

0.086

0.088

0.09

0.092

0.094

5 1 0 1 5 2 0 2 5 3 0 3 5 4 0

CA

RB

ON

DIO

XID

E EM

ISSI

ON

(K

G/M

W)

PRESSURE RATIO

ζ = 0%

ζ = 5%

ζ = 5%, ω=5%

77

4.11 Exergy destruction in combined cycle system

Figure 4.29 Percentage of exergy destruction in each component as compared to overall destruction in combined cycle with steam injection

Figure 4.28 shows exergy destruction in each individual component in combined

cycle compared to overall exergy destruction in the cycle. These results are obtained

using fixed turbine inlet temperature of 1200° C and pressure ratio of 25. The main

sources of exergy destruction in the combined cycle unit are the main combustion

chamber (CC1), reheat combustor (CC2) and heat recovery steam generator which

are responsible for 37 %, 20 %, and 16 % respectively of the total exergy destruction.

These results clearly show that the combustors of topping cycle have the highest

exergy destruction. In the other words, reducing the destruction in the combustors

of topping cycle will lead to a significant improvement in the exergetic efficiency and

also reduced destruction in the combined cycle. The exergy destruction in the

combustion champers is related to chemical reaction that occurs in combustion

process. The exergy destruction ratios associated with both turbines are less than

10 % of total exergy destruction of the power plant. Although the rejected heat in the

condenser is considered as tremendous amount from first law of thermodynamics

0

5

10

15

20

25

30

35

40

Exer

gy D

estr

uct

ion

(%

of

ove

rall

des

tru

ctio

n)

Combined cycle components

0 % steam injection

5 % steam injection

78

perspective, the exergy destruction ratio associated with the condenser unit is low

because the steam at condenser condition does not have potential power to produce

useful work. As fraction is steam (5 %) injected in CC1 and CC2, there is a drop of

exergy destruction in CC1 and CC2 by 2.3 % and 2 % respectively. Steam injection

adds more useful work and lowers the amount of the fuel in combustion chamber

that reduces the destruction.

4.12 Result validation

There are limited references with which to compare the new ideas research.

However, a partial validation with available literature and data is performed where

possible. For instance, effect of turbine inlet temperature on thermal efficiency has

been validated from results obtained by Kumar [31]. Result from research paper is

compared with present work in table 4.9. Not all combined cycle system components

in the research paper and present study are the same. Results are validated only

with limited range studied in present study. For example, in table 4.10, combined

cycle thermal efficiency validated only with pressure ratio range of 15 to 25. Other

than outside parameter are shown to understand result trend.

Table 4.9 Combined cycle thermal efficiency validation with reference [31]

Turbine Inlet Temperature (K)

Thermal efficiency (%)

[31]

Thermal efficiency (%) (present work)

1600 58 56

1700 59 57

1800 60 58

1900 61 59

Srinivas [15] studied combined cycle power generation extensively without steam

injection and with steam injection. Table 4.10 shows the comparison of thermal

efficiency for various pressure ratio between research paper and present study.

Initial pressure ratio of 10 and 15 shows slightly different when compared both

79

studies due to different assumption for ambient condition used in both studies.

However near optimum pressure ratio range of 20 to 35, there is a negligible

difference in thermal efficiency. Result shows in table 4.10 is valid only for

configuration 1 studied in present study.

Table 4.10 Combined cycle thermal efficiency validation with reference [15]

Pressure Ratio Thermal efficiency (%)

[15] Thermal efficiency (%) (present

study)

10 47.4 41

15 47.6 45

20 48 47

25 48.4 48.1

30 48.2 48.2

35 47.4 47.8

Table 4.11 shows result validation of exergy destruction in each major components

of combined cycle system compared with results obtained by Ahmadi [14] and

Srinivas [15]. Although not all components are similar to the research paper, major

components are common and results obtained from paper is within 2% of accuracy

with the results carried out in present study.

80

Table 4.11 Combined cycle component exergy destruction validation with references [14] and [15]

Combined cycle

components

Exergy destruction (%

of overall destruction)

[14]

Exergy destruction (%

of overall destruction)

[15]

Exergy destruction (%

of overall destruction)

(present work)

Air compressor stage 1

2 2 3

Air compressor stage 2

2 2 3

Combustion chamber

39 38 37

Gas turbine 5 2 4

Heat recovery steam generator

21 8 20

Steam turbine 6 4 7.5

Condenser 3 1 3.5

Overall exergy efficiency of combined system for configuration 2 is validated with

results obtained by Srinivas [15] shows in table 4.12. Although difference between

research paper and present study is 5%, there is a linear relationship between two

studies. Difference is mainly due to each components’ efficiency and state condition

used in reference study are different than present study.

Table 4.12 Combined cycle component exergy efficiency validation with reference [15]

Steam Injection

(%) Exergy efficiency

(%) [15]

Exergy efficiency (%) (Present work)

1 48 53.2

2 48.3 53.7

3 48.6 54.3

4 49 54.8

5 49.3 55.6

6 49.6 56.1

81

Chapter 5: Conclusions

This chapter summarizes the principal findings and the contributions from the

present work. Energy and exergy examinations of gas turbine combined cycle

(GTCC) configuration performed with and without steam injection. In addition,

present work investigates the combined cycle performance from changing operating

conditions such as turbine inlet temperature (TIT), ambient temperature, pressure

ratio and steam injection ratio. At the end provides some recommendations for the

future work.

5.1 Principal contributions

The compression ratios, air to fuel ratio as well as the isentropic efficiencies

are strongly influenced on the overall thermal and exergy efficiency of the

combined cycle gas turbine power plant. The overall thermal efficiency

increases and total power output increases linearly with the increase of the

compression ratio with constant turbine inlet temperature.

The steam injection increases gas cycle efficiency and decreases the steam

cycle efficiency. At fixed turbine inlet temperature of 1200° C, it is assumed

that complete combustion takes place in the GTCC and has excess air to

be used in reheater combustion chamber. Steam injection decreases the

amount of excess air in the combustion chamber as it also controls the

temperature. The air should not decrease below the minimum requirement

for complete combustion of the fuel. It implies that the amount of steam

injection has a limit depending on the air quantity in the compressor. From

the chemical reaction equation (3.36) in the gas reheater, (𝜙-2α-2β) is the

amount of excess oxygen. Value of 𝜙 should not be less than 4.

82

Steam injection in CC1 or CC2 individually has more benefit than steam

injection in both combustion chamber together. When steam is injected in

both combustion chambers, it lowers the flue gas temperature which has a

significant impact on lower steam production in bottom cycle.

In oxyfuel combustion, higher ratio of recycle flue gas brings higher thermal

efficiency. When steam is injected in CC1 only, highest thermal efficiency

(73.4%) achieved when 5% steam is injected. Adding more steam in

combustion chamber brings the thermal efficiency down.

5.2 Conclusions

The peak overall efficiency occurred at the higher compression ratio of 25

in combustion with air and oxyfuel combustion.

The flue gas from combustion chamber decreases with the increase in the

stem injection due to the decreased airflow rate. The steam entering into

the condenser decreases with the increase in the steam mass ratio.

Largest exergy destruction observed in combustion chamber (37% of

overall exergy destruction). Exergy destruction is lowered by 2% when 5%

steam is injected in CC1 and CC2.

Maximum steam to air mass flow should not be more than 10%. Injecting

more than 10% steam reduces oxygen contents in reheater combustion that

produces carbon monoxide instead of carbon dioxide.

Ideal fuel contribution between CC1 and CC2 is 54% and 46% respectively

Topping cycle work output increases with ambient temperature while steam

cycle work output decreases with corresponding ambient temperature.

Overall combined cycle work output increases with higher ambient

temperature as work output increase in topping cycle is higher than work

output decrease in steam cycle. Combined cycle workout increases with

steam injection.

83

5.3 Recommendations

The combined cycle power plant requires extensive studies to enhance their

performance. This study attempts to investigate several options to improve the

overall plant performance. The recommended future work is summarized as follows:

Present work focuses on the combined cycle power plant with steam

injection from thermodynamic point of views and the economic analysis was

not performed. Therefore, it would be worthwhile to study the combined

cycle power plant with steam injection from thermo-economic perspective

especially for the oxyfuel combustion cycle where cost to separate oxygen

from air has significant impact.

The combined cycle efficiency and work output with steam injection are

estimated based on the operating parameters of the bottoming cycle being

fixed. It would be worthwhile to investigate and optimize combined cycle

performance with varying the bottoming cycle operating parameters.

Present work concentrates on fuel being natural gas with 100% methane.

Study should be performed with various composition of natural gas as well

as biogas and syngas to understand the impact of steam injection on overall

plant performance.

Overall plant performance and emission characteristics should be studied

when fuel oxidizer is air in the primary combustion chamber and reheater

combustion chamber oxidizer is pure oxygen.

84

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[1] National Energy Board (Canada) “2009 Reference Case Scenario: Canadian Energy Demand and Supply to 2020” - An Energy Market Assessment. [2] “Ontario Gas Fired Generation Update”, Union Gas customer presentation, March 2011 [3] CO2 Emissions from Fuel Combustion (2009 Edition), International Energy Agency [IEA], Paris. [4] Canadell JG, et al.: Proceedings of the National Academy of Sciences, 104:18866, November 20, 2007 [5] Jacob, Daniel – “Introduction to Atmospheric Chemistry”, Princeton University Press. pp. 25–26. ISBN 0-691-00185-5, 1999 [6] [http://www.ec.gc.ca/doc/ed-es/p_123/CC_Plan_2007_e.pdf] [7] Nag P. K., “Power Plant Engineering” Tata McGraw-Hill Publication company Limited, New Delhi, Second Edition, 2004 [8] Franco A., and Casarosa C., “On Some Perspectives for Increasing The Efficiency of Combined Cycle Power Plants,” Applied Thermal Engineering, 22, pp. 1501-1518, 2002 [9] Dincer I., “The Role of Exergy In Energy Policy Making,” Energy Policy, 30, pp.137- 14, 2002 [10] Leo T. J., Grande I.P., and Notario P. P., “Gas Turbine Turbocharged by a Steam Turbine: A Gas Turbine Solution Increasing Combined Power Plant Efficiency and Power,” Applied Thermal Engineering, 23, pp. 1913-1929, 2003 [11] Sanjay Y., Singh O., and Prasad B.N., “Energy and Exergy Analysis of Steam Cooled Reheat Gas Steam Combined Cycle,” Applied Thermal Engineering, 27, pp. 2779-2790, 2007 [12] Khaliq A., and Kaushik S.C., Second-Law Based Thermodynamic Analysis of Brayton/Rankine Combined Power Cycle with Reheat, Applied Energy, 78, PP. 179-197, 2004

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[13] Butcher C.J., and Reddy B.V., “Second Law Analysis of a Waste Heat Recovery Based Power Generation System,” International Journal of Heat and Mass Transfer, 50, PP. 2355-2363, 2007 [14] Ameri M., Ahmadi P., and Khanmohammadi S., “Exergy Analysis of a 420 MW Combined Cycle Power Plant,” International Journal of Energy Research, 32, pp. 175-183, 2008 [15] T Srinivas, A V S S K S Gupta and B V Reddy – “Parametric simulation of steam injected gas turbine combined cycle” - Proceedings of the Institution of Mechanical Engineers, Part A: Journal of Power and Energy 2007 221: 873, 2007 [16] Horlock J.H., “Advanced gas turbine cycles”, Pergamon Press, Oxford edition, 2003 [17] Rolf Kehlhofer, Frank Hannemann, Franz Stirnimann, Bert Rukes – “Combined-Cycle Gas & Steam turbine Power Plants”- 3rd edition, PennWell Corporation, ISBN 978-1-59370-168-0, 2009 [18] Cengel Y.A., Boles M.A., “Thermodynamics An Engineering Approach”, McGraw-Hill, ISBN-10: 0073107689, 2006 [19] Marc A. Rosen, Ibrahim Dincer, Mehmet Kanoglu – “Role of exergy in increasing efficiency and sustainability and reducing environmental impact”, Energy Policy 36 128–137, 2008 [20] Ivar S. Ertesvag, Hanne M. Kvamsdal, Olav Bolland – “Exergy analysis of a gas-turbine combined-cycle power plant with precombustion CO2 capture”, Energy, Volume 30, Issue 1, P 5-39, ISSN 0360-5442, DOI: 10.1016/j.energy.2004.05.028, January 2005 [21] “Carbon Capture Journal”, ISSN 1757-2509, Issue 21, May 2011 [22] Cheng D.Y. US patent 4,297,841, 1981 [23] WLR Gallo, “A comparison between the HAT cycle and other gas-turbine based cycles: efficiency, specific power and water consumption” - Energy conversion and management, vol 38, no 15-17, pp 1595-1604, 1997 [24] J. De Ruyck, “ Efficient CO2 Capture through a Combined Steam and CO2 gas Turbine Cycle”, Energy conversion and management, Vol 33, No 5-8, pp 397-404, 1992

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87

Appendix: EES CODE

p_loss_ic=0 "Pressure loss in inter cooler" p_loss_cc=0 "Pressure loss in Combustion Chamber" p_loss_he_gas=0 "Pressure loss in Heat Exchanger - gas side" nc=.89 "Isentropic Efficiency of air compressors" ng=.89 "Isentropic Efficiency of Gas Turbines" nst=.80 "Steam Turbine efficiency" np=.95 "Pump Efficiency" p[1]=1 p[2]=(p[1]*p[4])^(1/2) p[3]=p[2]-((p[2]*p_loss_ic)/100) p[4]=PR*p[1] p[5]=p[4]-((p[4]*p_loss_cc)/100) p[6]=(p[5]*p[8])^(1/2) p[7]=p[6]-((p[6]*p_loss_cc)/100) p[8]=PR/p[5] p[9]=p[8]-((p[8]*p_loss_he_gas)/100) p[10]=100 p[11]=.1 p[12]=.1 p[13]=p[10] p[14]=PR+(PR*0.05) PR=25 T[1]=25 TS[2]=Temperature(Air,h=hs[2]) T[2]=Temperature(Air,h=h[2]) T[3]=T[2]-100 Ts[4]=Temperature(Air,h=hs[4]) T[4]=Temperature(Air,h=h[4]) T[5]=TIT //TIT=1400 //TIT2=1400 Ts[6]=Temperature(Air,h=hs[6]) T[6]=Temperature(Air,h=h[6]) T[7]=TIT2 Ts[8]=Temperature(Air,h=hs[8]) T[8]=Temperature(Air,h=h[8]) T[9]=85

88

T[10]=500 T[11]=Temperature(Water,P=P[11],h=h[11]) T[12]=T_sat(Water,P=p[12]) Ts[13]=Temperature(Water,x=0,h=hs[13]) T[13]=Temperature(Water,h=h[13],x=0) T[14]=Temperature(Steam,P=P[14],x=1) s[1]=Entropy('Air',p=p[1],h=h[1]) ss[2]=s[1] s[2]=Entropy(Air,T=T[2],P=P[2]) s[3]=Entropy(Air,T=T[3],P=p[3]) ss[4]=s[3] s[4]=Entropy(Air,T=T[4],P=p[4]) s[5]=Entropy(Air,T=T[5],P=p[5]) ss[6]=s[5] s[6]=Entropy(Air,h=h[6],P=p[7]) s[7]=Entropy(Air,T=T[7],P=p[7]) ss[8]=s[7] s[8]=Entropy(Air,h=h[8],P=p[8]) s[9]=entropy(Air,T=T[9],P=p[9]) s[10]=entropy(Steam,T=T[10],P=p[10]) ss[11]=s[10] sf[11]=entropy(water,P=P[11],x=0) xs[11]=(ss[11]-sf[11])/sfg[11] s[11]=Entropy(steam,T=T[11],h=h[11]) s[12]=Entropy(Water,P=P[12],x=0.1) ss[13]=s[12] s[13]=entropy(water,P=p[13],T=T[13]) s[14]=Entropy(Steam,P=P[14],h=h[14]) h[1]=Enthalpy(Air,T=T[1]) hs[2]=Enthalpy(Air,P=p[2],s=ss[2]) h[2]=h[1]+(hs[2]-h[1])/nc h[3]=Enthalpy(Air,T=T[3]) hs[4]=Enthalpy(Air,P=p[4],s=ss[4]) h[4]=h[3]+(hs[4]-h[3])/nc hs[6]=enthalpy(Air,P=p[6],s=ss[6]) h[6]=h[5]-(h[5]-hs[6])*ng hs[8]=enthalpy(Air,P=p[8],s=ss[8]) h[8]=h[7]-(h[7]-hs[8])*ng h[9]=enthalpy(Air,P=p[9],s=s[9]) h[10]=Enthalpy(Steam,T=T[10],P=P[10]) hf[11]=enthalpy(water,P=P[11],x=0)

89

hfg[11]=enthalpy(water,P=P[11],x=0.71) hs[11]=hf[11]+xs[11]*hfg[11] h[11]=h[10]-(h[10]-hs[11])*nst h[12]=Enthalpy(Water,P=P[12],x=0.1) hs[13]=Enthalpy(Water,P=p[13],s=ss[13]) h[13]=h[12]+(hs[13]-h[12])/np h[14]=Enthalpy(steam,P=P[14],x=1) hst=PR/(P[10]-P[11]) m_air[4]=m[1] m[1]=m[2] m[2]=m[3] m[3]=m[4] m[4]=(phi+(theta*3.76)+(4-lambda))*mfa mfa=1 m_fuel[4]=alpha m_steam[4]=zeta m[5]=m_air[4]+m_fuel[4]+m_steam[4] AF[1]=m_air[4]/m_fuel[4] m[6]=m[5] m_fuel[6]=beta m_steam[6]=omega m[7]=m[6]+m_fuel[6]+m_steam[6] m[8]=m[7] m[9]=m[8] m[10]=(m[8]*(h[8]-h[9]))/(h[10]-h[13]) m[11]=m[10]-m[14]-m[15] m[12]=m[11] m[13]=m[12] m[14]=zeta m[15]=omega alpha=al al=.8 beta=bet bet=.1 phi=(2*Theo_air/100) theta=(2*Theo_air/100) lambda=4 zeta=st*m[4] zeta=st*m_fuel[4] st=0 omega=st2*m[6] omega=st2*m_fuel[6]

90

st2=0 Theo_air =200 "% ASU=225/3600*phi/1000*32 W_c[1]=m_air[4]*(h[2]-h[1]) W_c[2]=m_air[4]*(h[4]-h[3]) W_com[1]=W_c[1]+W_c[2] W_GT[1]=m[5]*(h[5]-h[6]) W_GT[2]=m[7]*(h[7]-h[8]) W_GT_TOT[1]=W_GT[1]+W_GT[2] W_pump[1]=m[12]*(h[13]-h[12]) W_Top=W_GT_TOT[1]-W_com[1] W_ST[1]=(m[10]*(h[10]-h[11]))-(((m[14]+m[15])*(h[14]-h[11])))" W_net[1]=W_GT[1]+W_GT[2]+W_ST[1]-W_c[1]-W_c[2]-W_pump[1] X_fuel_alpha=(alpha)*836420 "kJ/kmol" X_fuel_beta=(beta)*836420 "kJ/kmol" LHV=(alpha+beta)*802303 ch44=-74873 co22=-393522 h2oo=-241827 eta_I[1]=(W_net[1]/((m[5]*(h[5]-h[4])+(m[7]*(h[7]-h[6])))))*100 eta_I[2]=(W_net[1]/(LHV))*100 eta_II[1]=(W_net[1]/(X_fuel_alpha+X_fuel_beta))*100 X_dot[1]=0 X_dot[2]=m_air[4]*((h[2]-h[1])-(T[1]*(s[2]-s[1]))) X_dot[3]=m_air[4]*((h[3]-h[1])-(T[1]*(s[3]-s[1]))) X_dot[4]=m_air[4]*((h[4]-h[1])-(T[1]*(s[4]-s[1]))) X_dot[5]=m[5]*((h[5]-h[1])-(T[1]*(s[5]-s[1]))) X_dot[6]=m[6]*((h[6]-h[1])-(T[1]*(s[6]-s[1]))) X_dot[7]=m[7]*((h[7]-h[1])-(T[1]*(s[7]-s[1]))) X_dot[8]=m[8]*((h[8]-h[1])-(T[1]*(s[8]-s[1]))) X_dot[9]=m[9]*((h[9]-h[1])-(T[1]*(s[9]-s[1]))) X_dot[10]=m[10]*((h[10]-h[1])-(T[1]*(s[10]-s[1]))) X_dot[11]=m[11]*((h[11]-h[1])-(T[1]*(s[11]-s[1]))) X_dot[12]=m[12]*((h[12]-h[1])-(T[1]*(s[12]-s[1]))) X_dot[13]=m[13]*((h[13]-h[1])-(T[1]*(s[13]-s[1]))) X_dot[14]=m[14]*((h[14]-h[1])-(T[1]*(s[14]-s[1]))) X_dot[15]=m[15]*((h[14]-h[1])-(T[1]*(s[14]-s[1]))) Ed_c[1]=X_dot[1]-X_dot[2]+W_c[1] Ed_c[2]=X_dot[3]-X_dot[4]+W_c[2]

91

Ed_c[5]=Ed_c[1]/Ed_tot*100 Ed_c[6]=Ed_c[2]/Ed_tot*100 Ed_GT[1]=X_dot[5]-X_dot[6]-W_GT[1] Ed_GT[2]=X_dot[7]-X_dot[8]-W_GT[2] GT[5]=Ed_GT[1]/Ed_tot*100 Ed_GT[6]=Ed_GT[2]/Ed_tot*100 Ed_HRSG[1]=X_dot[8]-X_dot[9]-X_dot[10]+X_dot[13] Ed_HRSG[5]=Ed_HRSG[1]/Ed_tot*100 Ed_ST[1]=X_dot[10]-(X_dot[11]+X_dot[14]+W_ST[1]) Ed_ST[5]=Ed_ST[1]/Ed_tot*100 Ed_CON[1]=X_dot[11]-X_dot[12] Ed_CON[5]=Ed_CON[1]/Ed_tot*100 Ed_pump[1]=X_dot[12]+W_pump[1]-X_dot[13] Ed_pump[5]=Ed_pump[1]/Ed_tot*100 Ed_cc[1]=X_dot[4]+X_fuel_alpha+X_dot[14]-X_dot[5] Ed_cc[5]=Ed_cc[1]/Ed_tot*100 Ed_cc[2]=X_dot[6]+X_fuel_beta+X_dot[15]-X_dot[7] Ed_cc[6]=Ed_cc[2]/Ed_tot*100 Ed_tot=Ed_c[1]+Ed_c[2]+Ed_GT[1]+Ed_GT[2]+Ed_HRSG[1]+Ed_ST[1]+Ed_CON[1]+Ed_pump[1]+Ed_cc[1]+Ed_cc[2] "Equation for Combustion Chamber CC1" HR1=(alpha*fuel)+(phi*O2)+(3.76*theta*N2)+(zeta*steam)+((4-lambda)*CO2) HP1=((alpha)*(CO2_formation+CO2_TIT-CO2_a))+(((2*alpha))*(H2O_formation+steam_TIT-H2O_a))+((phi-2*alpha)*(O2_TIT-O2_a))+(3.76*theta*(N2_TIT-N2_a))+zeta*steam_TIT+(4-lambda)*CO2_TIT //HP1=((alpha+1-lambda)*(CO2_TIT))+(((2*alpha+zeta))*(steam_TIT))+((phi-2*alpha)*(O2_TIT))+(3.76*theta*(N2_TIT)) HR1=HP1 "Equation for REHEATER CC2" HR2=((alpha+4-lambda)*CO2_reh)+((2*alpha+zeta)*steam_reh)+((phi-2*alpha)*O2_reh)+(3.76*theta*N2_reh)+((beta)*fuel)+(omega*steam) //HP2=((alpha+beta-lambda+1)*(CO2_formation+CO2_TIT2))+((omega+zeta+2*alpha+2*beta)*(H2O_formation+steam_TIT2))+((phi-2*alpha-2*beta)*(O2_TIT2-O2_a))+(3.76*theta*(N2_TIT2-N2_a)) //HP2=((beta-lambda+4)*(CO2_formation+CO2_TIT2))+((2*beta)*(H2O_formation+steam_TIT2))+((phi-2*alpha-2*beta)*(O2_TIT2-O2_a))+(3.76*theta*(N2_TIT2-

92

N2_a))+((alpha+4-lambda)*(CO2_TIT2))+((2*alpha+zeta)*steam_TIT2)+((phi-2*alpha)*O2_TIT2)+(omega*steam_TIT2) HP2=((beta)*(CO2_formation+CO2_TIT2))+((2*beta)*(H2O_formation+steam_TIT2))+((phi-2*alpha-2*beta)*(O2_TIT2-O2_a))+(3.76*theta*(N2_TIT2-N2_a))+((alpha+4-lambda)*(CO2_TIT2))+((2*alpha+zeta)*steam_TIT2)+((phi-2*alpha)*O2_TIT2)+(omega*steam_TIT2) HR2=HP2 CO2_Red=((alpha*16+beta*16)*2.75)/W_net[1]*1000 cp[1]=Cp(CO2,T=TIT) cp[2]=Cp(H2O,T=TIT) cp[3]=Cp(O2,T=TIT) cp[4]=Cp(N2,T=TIT) tot_mol[1]=(alpha+4-lambda)+(2*alpha+zeta)+(phi-2*alpha)+(3.76*theta) cp[5]=(((alpha+4-lambda)*(cp[1]))+(((2*alpha)+zeta)*(cp[2]))+((phi-2*alpha)*(cp[3]))+(3.76*theta*(cp[4])))/tot_mol[1] h[5]=cp[5]*TIT cp2[1]=Cp(CO2,T=TIT2) cp2[2]=Cp(H2O,T=TIT2) cp2[3]=Cp(O2,T=TIT2) cp2[4]=Cp(N2,T=TIT2) tot_mol[2]=(alpha+4-lambda)+(2*alpha+zeta)+(phi-2*alpha)+(theta*3.76)+beta+omega cp[7]=((alpha+beta-lambda+4)*(cp2[1])+(2*alpha+2*beta+zeta+omega)*(cp2[2])+(phi-2*alpha-2*beta)*(cp2[3])+(theta*3.76)*(cp2[4]))/tot_mol[2] h[7]=cp[7]*TIT2 "################# Entropy Generation CC1 ########################" m_r_cc[1]=alpha+phi+(theta*3.76)+zeta+(4-lambda) m_p_cc[1]=(alpha+4-lambda)+(2*alpha+zeta)+(phi-2*alpha)+(theta*3.76) s_r_cc[1]=alpha*Entropy(CH4,T=T[1],P=P[1])+phi*Entropy(O2,T=T[4],P=P[4])+(theta*3.76)*Entropy(N2,T=T[4],P=P[4])+zeta*Entropy(steam,T=TIT,P=P[14])+(4-lambda)*Entropy(CO2,T=T[4],P=P[4])+(8.31/273*(alpha*ln(alpha*P[4]/m_r_cc[1])+phi*ln(phi*P[4]/m_r_cc[1])+(theta*3.76)*ln(theta*3.76*P[4]/m_r_cc[1])+zeta*ln(zeta*P[4]/m_r_cc[1])+(4-lambda)*ln((4-lambda)*P[4]/m_r_cc[1])))

93

s_p_cc[1]=(alpha+4-lambda)*Entropy(CO2,T=T[5],P=P[5])+(2*alpha+zeta)*Entropy(H2O,T=T[5],P=P[5])+(phi-2*alpha)*Entropy(O2,T=T[5],P=P[5])+(theta*3.76)*Entropy(N2,T=T[5],P=P[5])+((8.31/273)*(((alpha+4-lambda)*ln((alpha+4-lambda)*P[4]/m_p_cc[1])+(2*alpha+zeta)*ln((2*alpha+zeta)*P[4]/m_p_cc[1])+(phi-2*alpha)*ln((phi-2*alpha)*P[4]/m_p_cc[1])+(theta*3.76)*ln(theta*3.76*P[4]/m_p_cc[1])))) s_cc[1]=s_p_cc[1]-s_r_cc[1] "################# Entropy Generation CC2 ########################" m_r_cc[2]=alpha+4-lambda+2*alpha+zeta+phi-2*alpha+(theta*3.76)+beta+omega m_p_cc[2]=(alpha+beta-lambda+4)+(2*alpha+2*beta+zeta+omega)+(phi-2*alpha-2*beta)+(theta*3.76) s_r_cc[2]=(alpha+4-lambda)*Entropy(CO2,T=T[6],P=P[6])+(2*alpha+zeta)*Entropy(H2O,T=T[6],P=P[6])+(phi-2*alpha)*Entropy(O2,T=T[6],P=P[6])+(theta*3.76)*Entropy(N2,T=T[6],P=P[6])+beta*Entropy(CH4,T=T[1],P=P[1])+omega*Entropy(H2O,T=T[14],P=P[14])+(8.31/273)*(((alpha+4-lambda)*ln((alpha+4-lambda)*P[6]/m_r_cc[2])+(phi-2*alpha)*ln((phi-2*alpha)*P[6]/m_r_cc[2])+(theta*3.76)*ln(theta*3.76*P[6]/m_r_cc[2])+beta*ln(beta*P[6]/m_r_cc[2])+omega*ln(omega*P[6]/m_r_cc[2]))) s_p_cc[2]=(alpha+beta+4-lambda)*Entropy(CO2,T=T[7],P=P[7])+(2*alpha+2*beta+zeta+omega)*Entropy(H2O,T=T[7],P=P[7])+(phi-2*alpha-2*beta)*Entropy(O2,T=T[7],P=P[7])+(theta*3.76)*Entropy(N2,T=T[7],P=P[7])+8.31/273*(((alpha+beta+4-lambda)*ln((alpha+beta+4-lambda)*P[7]/m_p_cc[2])+(2*alpha+2*beta+zeta+omega)*ln((2*(alpha+beta+zeta))*P[7]/m_p_cc[2])+(phi-2*alpha-2*beta)*ln((phi-2*alpha-2*beta)*P[7]/m_p_cc[2])+(theta*3.76)*ln(theta*3.76*P[7]/m_p_cc[2]))) s_cc[2]=s_p_cc[2]-s_r_cc[2] fuel=Enthalpy(CH4,T=T[1]) O2=enthalpy(O2,T=T[4]) N2=enthalpy(N2,T=T[4]) steam=Enthalpy(STEAM,T=T[14], P=P[5]) CO2=9364 H2O_formation=-241820

94

CO2_formation=-393520 CO2_a=9364 H2O_a=9904 O2_a=8682 N2_a=8669 O2_TIT=enthalpy(O2,T=TIT) N2_TIT=enthalpy(N2,T=TIT) steam_TIT=enthalpy(steam,T=TIT,P=P[5]) CO2_TIT=enthalpy(CarbonDioxide,T=TIT,P=P[5]) O2_reh=enthalpy(O2,T=T[6]) N2_reh=enthalpy(N2,T=T[6]) steam_reh=Enthalpy(steam,T=T[14], P=P[7]) CO2_reh=enthalpy(CarbonDioxide,T=T[6],P=P[6]) O2_TIT2=enthalpy(O2,T=TIT2) N2_TIT2=enthalpy(N2,T=TIT2) steam_TIT2=Enthalpy(Steam,T=TIT2,P=P[7]) CO2_TIT2=enthalpy(CarbonDioxide,T=TIT2,P=P[7])