exergy analysis of pulverized coal-fired ultra

EXERGY ANALYSIS OF PULVERIZED COAL-FIRED

ULTRA SUPERCRITICAL POWER PLANTS

A Thesis

Submitted to the Faculty of Graduate Studies and Research

In Partial Fulfillment of the Requirements

for the Degree of

Master of Applied Science

In Industrial Systems Engineering

University of Regina

by

Sandhya Hasti

Regina, Saskatchewan

August, 2013

Copyright 2013: Sandhya Hasti

UNIVERSITY OF REGINA

FACULTY OF GRADUATE STUDIES AND RESEARCH

SUPERVISORY AND EXAMINING COMMITTEE

Miss Sandhya Hasti, candidate for the degree of Master of Applied Science in Industrial Systems Engineering, has presented a thesis titled, Exergy Analysis of Pulverized Coal-Fired Ultra Supercritical Power Plants, in an oral examination held on Thursday, July 18, 2013.. The following committee members have found the thesis acceptable in form and content, and that the candidate demonstrated satisfactory knowledge of the subject material. External Examiner: Dr. Fanhua Zeng, Petroleum Systems Engineering

Supervisor: Dr. Adisorn Aroonwilas, Industrial Systems Engineering

Committee Member: *Dr. Raphael Idem, Industrial Systems Engineering

Committee Member: Dr. Amornvadee Veawab, Environmental Systems Engineering

Chair of Defense: Dr. Larena Hoeber, Faculty of Kinesiology and Health Studies *Not present at defense

i

Abstract

Demand for power generation has been increasing day by day and, along with it, a

clean and efficient way of generating power is becoming more and more important. Apart

from the electricity generated, the emission of the hazardous gases plays a vital role in the

research field. The world’s power generation majorly depends on fossil fuels like coal

and natural gases. Among power plants, pulverized coal-fired power plants are of great

concern due to the relatively high emission of CO2 per MW of production. The

modification of coal-fired power plants has been significantly used in reducing the

emissions of particulate matter, like CO2, to the atmosphere. Analysis of the power

generation systems is an expansive concept involving the efficient use of the energy

resources. In the present scenario, it is important to focus on the stability of the power

generation process. A study performed to analyze the stability of the system is known as

exergetic analysis. This research deals mainly with the exergy analysis of coal-fired

power plants operating in ultra supercritical conditions. The main steam conditions used

for the model development are 750oC temperature and 35 MPa pressure.

The primary objective of this study is to analyze the various working parametric

conditions that lead to the highest exergy destruction. Exergetic analysis is a

methodology for evaluation of the performance of components and involves examining

the exergy at different points in a series of energy-conversion steps. The power plant

model was validated using different operating conditions. The results showed consistent

performance for the comparison made. A parametric study was conducted for different

operating temperatures and pressures to determine the plant’s efficiency. The plant’s

overall thermal efficiency was determined to be 55.23% with a net efficiency of 44.16%

ii

for a gross power output of 941.25 MW. The results obtained from the exergy analysis

showed that the furnace had the highest exergy losses followed by turbine. The exergy

loss in the furnace was 886.75 MW and that of the turbine was 67.34 MW. The results

reveal that instability in the combustion process is the main reason for exergy loss in the

furnace. Instability is due to the greater entropy generation. Increasing the preheated air

temperature, maintaining the lowest possible moisture content in the coal, and decreasing

the excess air percentage decreased the exergy destruction rate in the furnace. A detailed

parametric study for the turbine to increase the exergetic efficiency was also conducted. It

was observed that operating the power plant at ultra supercritical conditions yields a

positive response regarding exergy losses.

iii

ACKNOWLEDGEMENTS

I would like to sincerely thank my professor, Dr. Adisorn Aroonwilas, for his

enormous support and valuable time throughout my study. His constant willingness to

mentor and motivate throughout my study inspired me to complete this work. I have

showed fullest dedication in this work, however, it would not have been possible without

the kind support of my professor. I also take this opportunity to show my gratitude to Dr.

Amornvadee Veawab for her valuable guidance. I sincerely thank her for providing a

good moral support.

Most importantly, I would like to acknowledge the Faculty of Graduate Studies

and Research (FGSR) at the University of Regina and the Natural Sciences and

Engineering Research Council of Canada (NSERC) for their financial support.

This work would not have been possible without the wishes of my parents and my

brother Mr. Suresh Hasti. I’m greatly indebted to thank my fiance Mr. Vignesh Ravi for

his personal support and encouragement all these years. Heartfelt thanks for my beloved

friends Ms. Amrutha, Mr. Prash and friends from Regina and India for being a wonderful

source of inspiration.

iv

TABLE OF CONTENTS

Page

ABSTRACT i

ACKNOWLEDGEMENT iii

TABLE OF CONTENTS iv

LIST OF TABLES vii

LIST OF FIGURES viii

NOMENCLATURE x

1. INTRODUCTION 1

1.1 Background 1

1.2 Electricity Generation Technologies 3

1.3 Availability of Resources 4

1.4 Pulverized Coal-fired Power Plants and its Performance 5

1.5 Need for Exergy Analysis 11

1.6 Research Objective 11

1.7 Thesis Outline 12

2. LITERATURE REVIEW AND FUNDAMENTALS 13

2.1 Coal Combustion 13

2.2 Steam Power Cycle and Energy Analysis 15

2.3 Exergy Analysis 21

2.4 Literature Review on Exergy of Power Plants 25

v

3. MODEL DEVELOPMENT 28

3.1 Integrated furnace System 28

3.1.1 Flue Gas Composition 28

3.1.2 Combustion Temperature 30

3.1.3 Air Preheater 32

3.1.4 Boiler section 33

3.2 Turbines and Pumps 36

3.3 Condenser 38

3.4 Feedwater Heaters 38

3.5 Computational Algorithm for Power Plant Model 39

3.6 Exergy Analysis for Individual Modules 44

3.6.1 Exergy Analysis of Integrated furnace System 44

3.6.2 Exergy Analysis of Turbines and Pumps 45

3.6.3 Exergy Analysis of Condenser 46

3.6.4 Feedwater Heaters 46

3.7 Overall Exergetic Efficiency of the Power Plant 47

3.8 Computational Model for Exergy Analysis 47

3.9 Power Plant Model Validation 52

4. RESULTS AND DISCUSSIONS 59

4.1 Base Performance of Ultra Supercritical Power Plant 59

4.2 Parametric Analysis of Ultra Supercritical Power Plant 68

4.2.1 Effect of Moisture Content in the Coal 68

vi

4.2.2 Effect of Excess Air Percentage 74

4.2.3 Effect of Air Preheated Temperature 77

4.2.4 Effect of Reheating Temperature 78

4.2.5 Effect of IP Turbine’s Inlet Pressure 82

4.2.6 Effect of LP Turbine’s Inlet Pressure 83

4.2.7 Effect of LP Turbine’s Exit Pressure 83

4.3 Optimum Operating Conditions 86

5 CONCLUSIONS AND FUTURE WORK 91

REFERENCES

APPENDIX

vii

LIST OF TABLES

Table 3.1 List of reference points used for development of power plant 33

Table 3.2 Enthalpy balance: formula for power plant model 42

Table 3.3 Exergy destruction: formula for power plant model 49

Table 3.4 Exergetic efficiency: formula for power plant 50

Table 3.5 Power plant model validation- Input information from literature 54

Table 3.6 Power plant model validation case analysis 55

Table 3.7 Exergy analysis validation–Literature operating parameters used 56

Table 3.8 Exergy model validation case analysis - Exergetic destruction 57

comparison from Wang et al., 2012

Table 3.9 Exergy model validation case analysis - Exergetic destruction 58

comparison from Aljundi, 2009

Table 4.1 Input process parameters used for simulation of base power plant 62

Table 4.2 Base performance of ultra supercritical coal-fired power plant 63

Table 4.3 Exergy destruction rate and exergy destruction percent 64

of the plant

Table 4.4 Optimal process parameters for ultra supercritical power plant 88

Table 4.5 Optimal results obtained for ultra supercritical power plant 89

viii

LIST OF FIGURES

Figure 1.1 Worldwide electricity demands through the year 2000-2030 2

Figure 1.2 Scheme of subcritical pulverized coal-fired power plant 8

Figure 1.3 Scheme of supercritical pulverized coal-fired power plant 9

Figure 1.4 Scheme of ultra supercritical pulverized coal-fired power plant 10

Figure 2.1 Scheme of steam power cycle 16

Figure 2.2 Scheme of pulverized coal-fired power plant 19

Figure 2.3 Scheme of furnace explaining the exergy balance 23

Figure 3.1 Scheme of steam flow inside furnace system 31

Figure 3.2 Scheme of pulverized coal-fired power plant with reference points 34

Figure 3.3 Algorithm for ultra supercritical coal-fired power plant model 43

Figure 3.4 Algorithm for exergy analysis of power plant model 51

Figure 4.1 Distribution of exergy destruction for ultra supercritical plant 65

Figure 4.2 Distribution of exergy rate inside furnace 66

Figure 4.3 Distribution of exergy percentage flow for ultra supercritical plant 67

Figure 4.4 Effect of moisture content at 5% excess air 69





Figure 4.9 Effect of excess air percent at 200oC preheated air temperature 75

Figure 4.10 Effect of excess air percent at 350oC preheated air temperature 76

Figure 4.11 Effect of reheating temperature at 0.7 MPa LPT inlet pressure 79

ix



Figure 4.14 Effect of IP inlet pressure with varying LP inlet pressure 84

Figure 4.15 Effect of IP exit pressure with varying LP inlet pressure 85

Figure 4.16 Effect of IP exit pressure with varying LP inlet pressure 85

Figure 4.17 Distribution of exergy percentage flow for optimized 90

ultra supercritical plant

x

NOMENCLATURE

fCp, Molar specific heat of flue gas, kJ/kmole.oC

Cp Heat capacity, kJ/kg.oC

Cp,a Heat capacity of the air, kJ/kg.oC

iCp , Molar specific heat of individual combustion product, kJ/kmole.oC

C Percent of carbon by weight, %

CO Moles of carbon monoxide, kmoles

CO2 Moles of carbon dioxide, kmoles

Hg Moles of mercury, kmoles

HHV High heating value, kJ/kg

H Percent of hydrogen by weight, %

H2O Moles of hydrogen dioxide, kmoles

hi Specific enthalpy, kJ/kg

inh Enthalpy of stream entering in, kJ/kg

outh Enthalpy of stream leaving out, kJ/kg

0h Enthalpy at the reference state, kJ/kg

1h Enthalpy of fluegas entering furnace, kJ/kg

2h Enthalpy of fluegas leaving furnace, kJ/kg

airPreheatedh Enthalpy produced by preheated air, kJ/kg

oal ch Enthalpy produced by coal, kJ/kg

fw h Specific enthalpy of water, kJ/kg

msatdrystea h Specific enthalpy of the saturated dry steam, kJ/kg

xi

satliquidh Specific enthalpy of the saturated liquid, kJ/kg

Evap h Specific enthalpy absorbed by steam in the evaporator, kJ/kg

steamh Specific enthalpy of superheated steam, kJ/kg

steamsatdryh Specific enthalpy of saturated dry steam, kJ/kg

idealout,h Ideal enthalpy at the turbine outlet, kJ/kg

actu a lou t,h Actual enthalpy for each turbine section, kJ/kg

outc ndh , Enthalpy of condensed steam at outlet, kJ/kg

incndh , Enthalpy of condensed steam at inlet, kJ/kg

inw aterh , Enthalpy of cooling water at inlet, kJ/kg

outw aterh , Enthalpy of cooling water at outlet, kJ/kg

2fw h Enthalpy of feedwater leaving economizer, kJ/kg

1fw h Enthalpy of feedwater entering economizer, kJ/kg

3fw h Enthalpy of the steam leaving evaporator, kJ/kg

steaminh Enthalpy of steam in, kJ/kg

steamouth Enthalpy of steam out, kJ/kg

LHV Low heating value of the coal, kJ/kg

L Latent heat of water vaporization, kJ/kg

M Percent of moisture in coal by weight, %

.

m Mass flow rate, kg/s

.

gm Mass flow rate of gas, kg/s

c

.m Mass flow rate of coal, kg/s

xii

cw

.m Mass flow rate of cooling water, kg/s

w

.m Mass flow rate of feedwater, kg/s

m.

a Mass of the air, kg/s

inm.

Mass of the entering stream , kg/s

outm.

Mass of the exiting stream, kg/s

m.

s Mass flow rate of the steam, kg/s

NOx Moles of nitrogen oxide, kmoles

NO2 Moles of nitrogen dioxide, kmoles

NS Entropy generation number

N O air , 2 Moles of oxygen in air, kmoles

i, 2ON Amount of oxygen required for combustion reaction i, kmoles

O, fuelN Moles of elemental oxygen in fuel, kmoles

excess , 2ON Amount of excess oxygen required, kmoles

total, 2ON Total moles of oxygen required for combustion, kmoles

NN 2 , remaining Amount of nitrogen remaining in the furnace, kmoles

supply , 2ON Amount of oxygen entering furnace, kmoles

N Fluegas Total amount of combustion product, kmoles

iN Number of moles of gaseous component i, kmoles

2CON Number of moles of carbon dioxide after combustion, kmoles

xiii

OHN 2 Number of moles of hydrogen dioxide after combustion, kmoles

2SON Number of moles of sulfur dioxide after combustion, kmoles

2NON Number of moles of nitrogen dioxide after combustion, kmoles

O2 Percent of oxygen by moles, %

O Elemental oxygen, kmoles

Po Atmospheric pressure, MPa

outCPP , Pressure at exit of condenser pump, MPa

inC PP , Pressure at inlet of condenser pump, MPa

outBPP , Pressure at exit of boiler feed pump, MPa

inBPP , Pressure at inlet of boiler feed pump, MPa

qsurr Heat loss due to the surroundings, kJ/kg

Q

.

Heat transfer rate, kJ/s

Q in

.

Heat rate at the inlet, kJ/s

.

K Q Heat transfer rate of the surroundings, kJ/s

Q out

.

Heat rate at the outlet, kJ/s

.

QRH Heat absorbed by reheater, kJ/s

AHQ.

Heat absorbed by air preheater, kJ/s

BoilerQ.

Heat absorbed by boiler, kJ/s

Q EVP

.

Heat absorbed by evaporator, kJ/s

Q ECN

.

Heat absorbed by economizer, kJ/s

xiv

.

QSH Heat absorbed by superheater, kJ/s

FurnaceQ.

Heat absorbed by furnace, kJ/s

SOx Moles of sulfur oxide, kmoles

SO2 Moles of sulfur dioxide, kmoles

sin Entropy of stream entering in, kJ/kg K

so Entropy at the reference state, kJ/kg K

sout Entropy of exiting stream, kJ/kg K

s1 Entropy of fluegas entering furnace, kJ/kg K

s2 Entropy of fluegas exiting furance, kJ/kg K

sgen

.

Entropy rate generated inside the system, kJ/kg K

s Specific entropy, kJ/kg K

S Percent of sulfur in weight basis, %

T Temperature, oC

Tk Surrounding temperature, oC

Ti Temperature of the system, oC

T0 Temperature of the reference property, oC

fT Temperature of the feed, oC

,fT 3 Temperature of flue gas leaving the combustion chamber, oC

,fT 4 Temperature of flue gas leaving the superheater, oC

,fT 5 Temperature of flue gas leaving the reheater, oC

,fT 6 Temperature of flue gas leaving the evaporator, oC

,fT 7 Temperature of flue gas leaving the economizer, oC

xv

∆TAH Temperature difference between inlet air and the preheated air, oC

Wa Amount of air supplied, kg of air/kg of coal

W

.

Work rate, kJ/s

.

W T,total Work done by turbine, kJ/s

W

.

HP Work done by high pressure turbine, kJ/s

W

.

IP Work done by intermediate pressure turbine, kJ/s

W

.

LP Work done by low pressure turbine, kJ/s

W

.

Pump Work done by pump, kJ/s

W

.

CP Work done by condensate pump, kJ/s

W

.

BP Work done by boiler feed pump, kJ/s

W

.

Net Net power output, kJ/s

W

.

out Work out from the furnace , kJ/s

d

.

X Exergy destroyed, kJ/s

AHd,

.

X Exergy destroyed in air preheater, kJ/s

DEd,

.

X Exergy destroyed in deaerator, kJ/s

CNDd,

.

X Exergy destroyed in condenser, kJ/s

CPd,

.

X Exergy destroyed in condensate pump, kJ/s

BPd,

.

X Exergy destroyed in condensate pump, kJ/s

xvi

Furnaced,

.

X Exergy destroyed in furnace, kJ/s

Boilerd,

.

X Exergy destroyed in boiler, kJ/s

IPd,

.

X Exergy destroyed in intermediate pressure turbine, kJ/s

LPd,

.

X Exergy destroyed in low pressure turbine, kJ/s

HPd,

.

X Exergy destroyed in high pressure turbine, kJ/s

Hd,

.

X Exergy destroyed in feedwater heaters, kJ/s

Td,

.

X Exergy destroyed in turbine, kJ/s

yi Mole fraction of component i

ya Specific humidity of air, kg of water vapor/kg of air

Greek letters

th Thermal efficiency of turbine, %

Efficiency, %

isentropic Isentropic efficiency, %

HHVNet, Net efficiency of turbine based on HHV of coal, %

LHVNet, Net efficiency of turbine based on LHV of coal, %

Ψ,Furnace Exergetic efficiency of the furance, %

Ψ,Turbine Exergetic efficiency of the turbine, %

Ψ,CP Exergetic efficiency of the condensate pump, %

xvii

Ψ,CND Exergetic efficiency of the boiler feed pump, %

Ψ,DE Exergetic efficiency of the deaerator, %

Ψ,AH Exergetic efficiency of the air preheater, %

Ψ,H Exergetic efficiency of the feedwater heater, %

Ψ,Boiler Exergetic efficiency of the boiler, %

Ψ,EVP Exergetic efficiency of the evaporator, %

Ψ,ECN Exergetic efficiency of the economizer, %

Ψ,SH Exergetic efficiency of the superheater, %

Ψ,RH Exergetic efficiency of the reheater, %

Ψ,total Exergetic efficiency of the overall power plant, %

Ψ Specific exergy, kJ/kg

inΨ Specific exergy in, kJ/kg

outΨ Specific exergy out, kJ/kg

cΨ Exergy content in the coal, kJ/kg

aΨ Exergy content in the air, kJ/kg

1 Specific volume of condensate liquid, m3/kg

2 Specific volume of boiler feed liquid, m3/kg

xviii

Acronym and Abbreviation

AH Air preheater

BP Boiler feed pump

CND Condenser

CP Condensate pump

DE Deaerator

ECN Economizer

EVP Evaporator

Eff Efficiency

G Generator

H Feedwater heater

HPH High pressure feed water heater

HP High pressure turbine

IP Intermediate pressure turbine

LP Low pressure turbine

LPH Low pressure feed water heater

RH Reheater

SH Superheater

USC Ultra supercritical coal-fired power plant

1

1. INTRODUCTION

1.1 Background

The demand for power generation has been increasing constantly over recent

decades (CEA, 2012), and along with it, an environmentally friendly and efficient way of

generating power has become more important. The demand for the electricity has been

increasing worldwide at a rapid rate as detailed in Figure 1.1 (IAEA, 2004). Today, world

power generation depends heavily on the combustion of fossil fuels such as coal and

natural gas (EIA, 2013). Coal is considered one of the important fuels among the other

fuels such as natural gas, oil, etc, due to the fact that it is abundant and less expensive

(EIA, 2012). Demand for a more efficient coal fired power plant has been increasing due

to the increase in the demand for electricity. Energy losses associated with the electricity

generation is of greater concern due to which the net efficiency decreases. Due to this fact,

the coal-fired power plant has been studied in a number of research projects with the

primary focus on the alternative design and operation of the plant to increase the net

efficieny . Analysis of the coal-fired power plant is a broad concept involving the efficient

use of energy resources. In earlier days, the energy efficiency of the power plant was

analyzed based on the first law of thermodynamics (Smith et al., 1996). However, in

recent times, the second law of thermodynamics has been widely used to determine

exergy losses to study the quality of the energy produced within the system in a broader

spectrum. This approach is commonly known as exergy analysis.

2

0

10000

20000

30000

40000

2000 2010 2020 2030

Year

Ele

ctri

city

Gro

ss G

en

era

tio

n (

Tril

lio

n k

ilo

wat

t H

ou

r)

Figure 1.1 Worldwide electricity demand through the year 2000-2030.

(Redrawn from IAEA, 2004)

3

1.2 Electricity Generation Technologies

A power plant is defined as the assembly of equipment that generates a flow of

mechanical or electrical energy (Raja et al., 2006). The equipment used is known as the

generator. Power plants are generally classified into two types: conventional power plants

and non-conventional power plants, and they are classified based on the electricity

generation devices and fuel type (Kitto, 1996).

Power plants classified regarding the electricity generation devices such as

turbines are called conventional power plants. Examples of conventional sources of

energy include coal, natural gas, petroleum, and water power. The device that drives

electricity generation determines the kind of power plant. For instance, steam turbine

plants use the dynamic pressure generated by expanding steam to run the blades of a

turbine. Some other kinds of the conventional power plants are: gas turbine plants,

combined cycle plants, internal combustion plants, pulverized coal-fired power plants,

circulating fluidized bed power plants, pressurized fluidized bed power plants, integrated

gasification cycle power plants, hydro-electric power plants, nuclear power plants, diesel

power plants, steam turbines, and steam engines (Michael et al., 2011). Among the above

different kinds of power plants, steam turbines, steam engines, diesel power plants, and

nuclear power plants are categorized as the thermal power plants because they convert

heat into the electric energy.

Power Plants that are classified based on fuel type are called non-conventional

power plants. Some of the fuels used are biomass, solar, biogas, wind, tidal, and

geothermal. Some examples of the power plants are thermo-electric generator, fuel cell

power plants, photovoltaic solar cell power systems, fusion reactors, geothermal energy

4

plants, wind energy power systems, tidal wave plants, and biogas and biomass energy

power systems.

1.3 Availability of Resources

The resources can be classified into two categories: renewable resources and non-

renewable resources. Fossil fuels are classified as non-renewable resources. Most of the

fossil fuel plants run using coal, oil, or gas fuels. The usage of these fuels depends upon

on the availability of the fuel. The percentage of fuel usage is coal by 60 percent, oil by 10

percent and gas by 30 percent (CEA, 2011b). Among all the available sources of fuel, oil

is considered to be the most efficient, and the calorific value of oil is very high when

compared to other fuels (WRI, 2007). However, this was the case 30 years ago before the

usage of oil was deliberately dropped to 10 percent from 30 percent (EIA, 2011). The

main reason behind the drop was the availability and price of the fuel. There has been

instability in the supply of oil (WRI, 2007). Because of these reasons, coal was considered

to be the most abundant fuel with the most reliable supply among all the fuels used. Also,

the price of oil is relatively high when compared to that of coal.

Coal was reported to be the least convenient fuel (CEA, 2011a) with the main

reason being the release of toxic substances to the environment due to the process of

combustion. Some of the harmful substances include ash, carbon dioxide (CO2), nitrogen

oxides (NOx), sulfur oxides (SOx), and mercury (Hg) (IEA, 2012). Coal is the primary

source of fuel for electricity generation in most of the countries, especially India, China,

Canada, and the United States (EIA, 2012). As high as 350,000 MW of electricity is

generated from coal-fired power plants in North America alone (U.S.DOE, 2005).

5

Because of the environmental concerns, the challenging part of using coal is to manage

the emissions of the harmful gases as a result of combustion process. The emission of

CO2, NOx, SOx, and Hg gases is very dangerous and harmful to the environment; in

particular, CO2 can lead to global warming.

To date, a number of technologies have been developed to reduce the emission of

CO2 including carbon capture and geological sequestration. It has been a challenge to

meet the “near zero emission” (NZE) criterion while using coal as the fuel in the power

plant.

1.4 Pulverized Coal-Fired Power Plants and its Performance

Pulverized coal-fired power plants were first discovered in the 1920s. Europe and

Asia led in the deployment of the most advanced pulverized coal systems, although the

plants were gaining renewed attention in North America as well. These power plants

supply 50 percent of the world’s power demand (WRI, 2012). They served as the main

backbone for the power industry in most countries. In recent decades, coal-fired power

plants have been used as the main mode of electricity generation, and they supply

electricity for over half of the countries in the world, such as Canada, Russia, China,

Indonesia, Australia, South Africa, and India (CEA, 2012). Over the past 80 years, the

increasing electricity generation has been derived from the use of coal-fired power plants

due to the simplicity in the power generation process and a very efficient conversion rate

of a little more than one-third of the fuel’s energy potential into power (IEA, 2012).

The process of pulverized coal-fired power generation begins by crushing coal

into a fine powder that is fed into a furnace system where it undergoes combustion to

6

produce heat. The produced steam is then used to drive one or more turbines to generate

electricity. Based on the operating temperature and pressure of the steam used, the power

plants are classified into subcritical, supercritical, and ultra supercritical power plants

(Smith et al., 1996). Thermodynamics is the conceptual study relating heat, work,

temperature, and energy (Michael et al., 2011). The process parameters, such as pressure

and temperature, determine the rate of energy flow or heat flow inside the system, and

these parameters are subjected to change along with change in the environmental

reference temperature and pressure. Equilibrium means that any given system is

physically at a “zero or dead state” meaning there is no net charge or net transfer of heat

between the systems (Perry et al., 1997). The physical changes inside any system can be

described using the zeroth, first, second, and third law of thermodynamics (Smith et al.,

1996).

A typical pulverized coal-fired subcritical power plant is shown in Figure 1.2.

These kinds of coal-fired power plants are operated below the critical pressure of water at

a main steam temperature of 538∘C and pressure of 16.54 MPa. The feedwater train

consists of six closed feedwater heaters (four low pressure and two high pressure) and one

open feedwater heater (deaerator). Extractions for the feedwater heaters, deaerator, and

the boiler feed pump were taken from the turbine. The overall plant net efficiency was

found to be around 37.6 percent (Marion et al., 2004; U.S.DOE, 1999; Lako, 2004). To

further increase the net efficiency of the power plant, it was necessary to raise the pressure

of the main steam used in the turbine beyond the critical condition of water, which is

around 22.06 MPa; such plants were called supercritical power plants. The flow diagram

of the supercritical power plant is shown in Figure 1.3. The design includes seven closed

7

feedwater heaters and one open feedwater heater. The main steam was maintained at a

higher pressure in the range of 24 MPa and temperature between 600∘C and 760∘C

(Leung and Moore, 1966) to boost the overall efficiency of the plant to at least 39.9

percent (U.S.DOE, 1999).

In recent times, there has been a great interest in the ultra supercritical power

plants operated at a very high temperature, 760∘C or above, and high pressure of 35 MPa.

The design of an ultra supercritical power plant is illustrated in Figure 1.4. The design

consists of eight closed feedwater heaters and an open feedwater heater. The increase in

the number of closed feedwater heaters helps in achieving a higher net efficiency rate of

approximately 42 percent (Kjaer, 2002).

8

HP High pressure turbine H1-H2 High pressure closed feedwater heaters BP Boiler feed pump IP Intermediate pressure turbine H3-H6 Low pressure closed feedwater heaters CP Condensate pump

LP Low pressure turbine DE Deaerator AH Air preheater

CND Condenser G Generator

Figure 1.2 Scheme of subcritical pulverized coal-fired power plant.

(Modified from Li and Liu, 2012)

9



CND Condenser G Generator

Figure 1.3 Scheme of supercritical pulverized coal-fired power plant.

(Modified from Bakhshesh and Vosough, 2012)

10



C Condenser G Generator

Figure 1.4 Scheme of ultra supercritical pulverized coal-fired power plant.


11

1.5 Need for Exergy Analysis

Carnot first detailed the concept of exergy analysis in the year 1824 followed by

Clasisus in the year 1865. The first law of thermodynamics (first law of analysis) is

applied to a system or process to determine the efficiency, whereas the second law of

analysis (exergy analysis) is based on the concept of irreversibility and entropy

production. Researchers are constantly applying the concept of the first law of analysis to

calculate the energy losses using the enthalpy balance over the coal-fired power (Gwosdz

et al., 2005; Kiga et al., 2000). In recent years, the concept of exergy analysis has gained

importance over the first law of analysis as the first law of thermodynamic analysis fails

to produce sufficient results to study the performance of a power plant. The objective of

exergy analysis is to identify the locations of exergy losses and to study the quality of the

power plant. Exergy analysis stands as a tool to locate the imperfections inside the process

or system, which first law of analysis fails to do. Exergy analysis not only locates the

irreversibility inside the system, but it also helps in assessing the efficiency of the

individual components (Kaushik et al., 2011). Findings from exergy analysis can help

process practitioners refine and develop the optimal coal-fired power plant.

1.6 Research Objective

This study aimed at developing a mathematical process model for an ultra

supercritical pulverized coal-fired power plant. The model is needed for investigating the

different parameters that will affect the overall performance of the power plant. The

investigation was mainly focused on on various process analogues that could improve the

net efficiency of the power plant. The primary task of this study was to develop a power

12

plant model based on the concepts of coal combustion, energy balances, enthalpy

changes, entropy changes, and heat transfer of the steam power cycle. The model

developed was a Macro-supported file using Visual Basic as the back end source to

determine thermodynamic properties including enthalpy and entropy of steam. After

model development, the model was simulated and its results were compared with data in

the literature so as to validate the model.

After the model validation, an exergy analysis for the power plant model was

carried out using the concept of the second law of thermodynamics. The analysis results

gave the overall exergetic performance of the plant as well as the information to identify

the component that has the highest exergy loss in the entire power plant model and to

come up with a solution to reduce the exergy loss for the developed ultra supercritical

pulverized coal-fired power plant.

1.7 Thesis Outline

This thesis is divided into five chapters. Chapter One presents the background on

pulverized coal-fired power plants and exergy analysis. Chapter Two contains the

literature review on coal-fired power plants and basic concepts of combustion steam

power cycle and exergy destruction. Chapter Three provides details of development and

validation of the ultra supercritical coal-fired power plant model. Chapter Three also

introduces the formulae used for the power plant model and exergy analysis in detail.

Simulation results for the power plant model as well as exergy analysis are discussed in

Chapter Four. Finally, conclusions drawn from the study and recommendations for future

work are given in Chapter Five.

13

2. LITERATURE REVIEW AND FUNDAMENTALS

This chapter describes the principles behind the coal combustion process, steam

power cycle, and the concept of second law analysis in detail. Second law analysis can

also be termed Exergy analysis. The literature available for pulverized coal-fired power

plants and exergy analysis are discussed herein.

2.1 Coal Combustion

The combustion process is defined as the release of heat energy from exothermic

chemical reaction by burning of fuel (Smith et al., 1996). As the combustion process of

coal is the reason for rating it as the least efficient fuel (Singer, 1991), it is important to

concentrate on reducing the inefficiencies caused by the combustion.

The combustion process liberates steam inside the furnace system. The liberated

steam is used to drive the series of turbines operated at different pressures. In the process

of steam transformation, the heat generated from the combustion converts the working

fluid (liquid water) entering from the feedwater heaters to the superheated steam, low

pressure steam, and high pressure steam based on the system requirements. As the series

of turbines rotates, it generates power from the generator. The basic chemical reaction

between oxygen in air with carbon (C), hydrogen (H), nitrogen (N) and sulfur (S) in coal

are given below in Reactions 2.1 through 2.4 (Perry et al., 1997).

14

CO2 CO2 (2.1)

SO2 SO2 (2.2)

NO2 NO2 (2.3)

H 1

4O2

1

2H2O (2.4)

It should be noted that the typical furnace requires an excess amount of air or

oxygen to ensure the complete combustion of coal. The flue gas composition of the

combustion products can be determined by performing material balance for Reactions 2.1

through 2.4 together with the known amount of supply air.

The amount of heat energy released from Reactions 2.1 through 2.4 determines

the heat of combustion, which can be expressed either in dry or wet mode based on the

moisture content in the fuel. The dry basis of heat of combustion is known as “High

Heating Value”, HHV (Smith et al., 1996). A small amount of this heat will be consumed

by water vaporization during the combustion process resulting in a lower amount of heat

energy available for the vapor power cycle. This reduced heat of combustion is known as

“Low Heating Value (LHV)”. The Equation to find the high heating value of coal is

(Perry et al., 1997):

HHV = 2.326 . [146.58C + 568.78H + 29.4S - 6.58A - 51.53 (O+N)] (2.5)

where C, H, S, O, N is the weight percentage of carbon, hydrogen, sulfur, oxygen, and

nitrogen, respectively. The equation to find the low heating value of coal is (Perry et al.,

1997):

LHV = HHV – (L .W) (2.6)

15

where L is the latent heat of water vaporization and W is the amount of water vapor in the

resulting flue gas per unit mass of coal burned. The equation to find the W is (Perry et al.,

1997):

W = M + (9. H) + (Ya .Wa) (2.7)

where M and H are the mass fractions of moisture and hydrogen, respectively. (Ya) is the

specific humidity of the air, and (Wa) is the actual amount of air supplied per kg of coal.

2.2 Steam Power Cycle and Energy Analysis

The operation of power plants and their performance based on the operating

conditions are well documented in the literature (Oktay, 2009; Kakaras et al., 2002;

U.S.DOE, 1999). A simple furnace system is shown in Figure 2.1 to illustrate the concept

of material and energy balances. Irrespective of the operating conditions inside the

process, the mass balance for the furnace section is given by the Equation (2.8):

in

.m -

m.

out = 0 (2.8)

where

m.

in and

m.

out denote the mass flow rate of water entering and leaving the system,

respectively. The heat transfer for each furnace section Furance

.Q can be written as

)h(hmQ ino u t in

.

Fu rn a ce .

(2.9)

where in

.m represents the mass flow rate of stream entering in, and hout and hin denote the

outlet and inlet specific enthalpy (Singer, 1991).

16

Figure 2.1 Scheme of steam power cycle.

(Modified from Smith et al., 1996)

H2O- in

Mass in = .

m in

Enthalpy in = hin

Entropy in = sin

H2O- out

Mass out = out

.m

Enthalpy out = hout

Entropy out = sout

Heat in

(.

Q Boiler)

From hot flue

gas

Turbine

Feedwater

Condenser

Pump

Superheater

Reheater

Evaporator

Economizer

17

William Rankine first introduced the Rankine cycle in 1859 (Singer, 1991). The

Rankine cycle starts from the boiler where superheated steam is generated from the

combustion heat. The liberated superheated steam drives the series of turbines to generate

electricity. A low pressure steam from the exit of the turbine is condensed into saturated

liquid water in a condenser. Heat rejection occurs during the condensation process where

the pressure is maintained very low and at a constant low temperature close enough to

ambient conditions (Smith et al., 1996). The pressure of the condensed liquid water is

raised by means of a boiler pump, and the liquid is sent back to the furnace to complete

the steam cycle. The drawback of the Rankine cycle is that the efficiency of the power

plant is rather low.

One way of increasing the efficiency of the power plant is by modifying the

Rankine cycle to a Reheat-Regenerative cycle, where a part of the steam is extracted

from the turbine and reheated inside the furnace to further increase the temperature of the

steam. This reheated steam drives the turbine harder to generate more power. As a result,

the net power output and the thermal efficiency of the power plant will increase. Another

way of increasing the efficiency is by incorporating feedwater heaters to further heat the

saturated liquid before it enters the furnace (Jayamaha, 2008). The condensed liquid is

fed into a series of open and closed feedwater heaters. The heating medium for the

feedwater heaters is steam extracted from the turbine. The amount of steam extracted to

heat the feedwater heaters is relatively small. In most of the steam-driven power plants, it

is proven that superheating and reheating the steam helps to improve the overall

performance of the steam turbine (Sue and Chang, 2004). Superheaters and reheaters help

in resolving the inefficiency problems (Perry et al., 1997). During the process of steam

18

expansion inside the turbine, the pressure reduces from a higher pressure to lower

pressure, and the process is said to be isentropic because the expansion takes place in a

reversible adiabatic manner. The compression process is also isentropic, although the rise

in the temperature is considerably less, but the temperature has to be increased. Adding

heat externally could help in achieving the higher temperature. The efficiency of the

power cycle depends on the boiler temperature, reheat temperature, percentage of excess

air used, and air preheater temperature. The air preheater is used to increase the

temperature of the input air from room temperature to 200 – 350∘C (Woodruff et al.,

2005). A typical flow diagram of a pulverized coal-fired power plant is shown in Figure

2.2, and the net power output from the turbine .

W T,total can be calculated by

.

W T,total = .

W HP +

.

W IP +

.

W LP (2.10)

where .

W HP,

.

W IP,

.

W LP are the power output produced from high pressure, intermediate

pressure, and low pressure turbines, respectively. Total pumping power input can be

written as

.

W Pump = .

W CP + .

W BP (2.11)

where .

W CP, .

W BP represent power input for the condenser feed pump and boiler feed

pump, respectively. The total power output from the power plant can be written as

.

W Net = .

W T,total – .

W Pump (2.12)

19

HP High pressure turbine IP Intermediate pressure turbine LP Low pressure turbine CP Condensate pump BP Boiler feed pump DE Deaerator

AH Air preheater CND Condenser G Generator

EVP Evaporator ECN Economizer SH Superheater H1-H7 Feedwater heaters E1-H7 Expansion valves RH Reheater

Figure 2.2 Scheme of pulverized coal-fired power plant.

(Modified from Li and Liu, 2012)

Feed water

BP

E1 E2 E4 E5 E6 E7

E3

AH

Coal

Air

SH RH

EVP

ECN

DE

H1 H2 H3 H4 H5 H6 H7

HP IP LP

Saturated Steam

Reheat Steam

CND

CP

Condensed Liquid

Furnace

G

20

Identifying the performance and irreversibility inside the system is very important.

The first second laws of thermodynamics play equally major roles in determining the

quality of the system. A research study performed by Cheng et al., (2010) details the

thermodynamic analysis of the power generation system, and it was demonstrated that

both the first second laws of thermodynamics helped to accomplish the quality analysis,

but most of the time, they are not recognized and clearly defined. Oktay (2009)

performed a case study that focused on a power plant located in Turkey. A comparative

analysis was made between fluidized bed power plants and conventional power plants.

The power plant was analyzed on the basis of exergetic and energetic performance.

Apart from performing thermodynamic analyses, it is important to perform

economic analyses to study the quality of the system. One such effort was made by

Casarosa et al. (2004) in which both thermodynamic and economic analyses were

performed to design the optimal heat recovery system generator. It was found that

performing an economic analyses for the steam power cycle could increase the efficiency

of the overall system. Wei et al. (2007) analyzed a typical Rankine cycle and proposed

numerous suggestions to improve the performance of the power cycle. One of the

improvement factors proposed was to cool the condenser properly with less heat loss so

that the efficiency of the condenser will be higher. Also, apart from concentrating on the

internal factors, it is important to concentrate on the external factors that affect the

performance of the power plant. Some of the external factors are environmental

conditions and boundary losses. They play a major role in determining the performance

of the power plant (Meyer et al., 2009).

21

2.3 Exergy Analysis

The thermodynamic analysis of any power plant considers the material, energy,

entropy, and exergy balances (Tsatsaronis, 2011). It is important to determine the amount

of work potential that can be attained from the system. The maximum work potential

derived from the system at any given reference temperature and pressure is often referred

to as “useful work or exergy” (Yunus and Michael, 2008). Since exergy deals with the

amount of useful work available in the system with respect to the reference point, exergy

is also termed as “availability or available energy” (Michael et al., 2011). The amount of

useful work produced depends on the conditions of the system and the immediate

surroundings outside the system. At steady state, the exergy balance equation for the

control volume system given in Figure 2.3 can be written as

0 = W - m - Q + m.

out

.

in

..

(2.13)

where ,

Q

.

, and

W

.

represent specific exergy, heat transfer rate, and work rate or

power, respectively (Kaushik et al., 2011). Entropy can be defined as a state variable

whose change is defined for a reversible process at a given temperature (Thess, 2011).

The entropy change relation for the control volume system depends on the inlet entropy

and outlet entropy difference.

Si n

.

i n

+ Qi n

T

.

i n

+ S.

gen = S.

out

out Qout

T

.

out

(2.14)

where ou tingenou tin QQSSS.

,

..

,

.

,

.

, and T represent the entropy rate at the inlet, entropy rate at

the outlet, entropy rate generated inside the system, heat rate at the inlet, heat rate at the

22

outlet, and temperature of the system (Aljundi, 2009). The entropy generation number

can be written as (Reddy et al., 2002):

gp

gens

mC

SN

..

.

(2.15)

where Ns is the entropy generation number, Cp is the specific heat value, and g

. m is the

mass flow rate of the gas. The general equation for the unit exergy can be written as

) s-(s T - )h-h ( =Ψ oo0 (2.16)

where h, s represent the specific enthalpy and specific entropy at temperature T and h0 , s0

and T0 are the specific enthalpy, specific entropy, and temperature of the reference

property (Regulagadda et al., 2010). The destruction of exergy in a system depends on

the entropy generation or changes happening within the system considered (Michael et

al., 2011). Whenever an entropy change is generated inside the system, a part of the

useful work that can be done by the system is destroyed, which indirectly leads to energy

loss, and this destruction of energy is known as Exergy destruction ( d

.

X ).

When the entropy generated is less, then the exergy destruction rate is also less,

and when the entropy generated is greater, the destruction is greater (Kaushik et al.,

2011). Thus, we can conclude that exergy destruction rate depends on entropy generation

rate. The general equation of exergy destruction for the system shown in the Figure 2.3

can be expressed as

genod

.S T =

.X . (2.17)

23

Figure 2.3 Scheme of furnace explaining the exergy balance

Environmental reference point

Ref Enthalpy in (ho)

Ref Entropy in (so)

Ref Temperature (To)

Ref Pressure (Po)

Exergy in ( inm.

in )

Enthalpy in (h1)

Entropy in (s1)

Exergy out ( outm.

out )

Enthalpy out (h2)

Entropy out (s2)

Heat in ( BoilerQ.

)

Work out (

.W out)

Furnace system

24

Exergy is developed based on the second law of thermodynamics where the analysis

depends on the entropy generation of the system. The entropy generation depends on the

inlet and the outlet entropy changes during the process.

)T

q + s-(s

.m =

.S

o

surrsgen 12 (2.18)

where

m.

s represent the flow rate of the steam entering the system and qsurr is the heat loss

due to the surroundings. For the control volume system at any given condition, the

general exergy balance equation is given by the relation below (Sue and Chuang, 2004):

out

...

d

..

in

..

in

k

.

Tk

T -1 Ψm X W = Ψm Q)( (2.19)

where k

.

Q represent the heat transfer rate of the surroundings and Tk is the surrounding

temperature. For any process, there are certain losses like heat loss and surrounding loss.

Heat loss is due to the heat and mass transfer inside the system and the losses that occur

due to the surroundings is known as surrounding losses (Thess, 2011). As a result of the

heat loss, there is loss in the work done.

25

2.4 Literature Review on Exergy of Power Plants

Lior and Zhang (2007) focused on defining the relation between energy, exergy,

and the second law of efficiency. Chaibakhsh and Ghaffari (2008) performed a detailed

study regarding the entropy generation that relates the heat and mass transfer concepts.

Nag and De (1997) performed a study to analyze the steam power cycle based on the

concepts of entropy generation number similar to Chaibakhsh’s work. Reddy et al. (2002)

also introduced an entropy generation number to find the entropy generation rate, but the

study made a comparative analysis of entropy generation with the help of non-

dimensionless parameters like heat capacity ratio and relative difference in the inlet gas

temperature ratio. Furthermore, an entropy generation equation similar to Equation 2.15

was derived to calculate the entropy generation inside the system.

For any system, efficiency determines the performance of the system. Improving

the system’s efficiency is considered important for any power plant. According to Suresh

et al. (2006), an efficient way to improve the power plant efficiency is by analyzing the

exergy destruction. Wang et al. (2012) described the exergy destruction as an irreversible

work that is mostly wasted, and the study illustrated the concept of irreversible work by

redefining the definition as the amount of work that can never be reversed or is wasted,

and this irreversibility in a system was found to reduce the plant’s overall performance.

In the literature, there are a number of research papers available to detail the

energetic and exergetic performance of coal-fired power plants. For instance, Habib and

Zubair (1992) performed an irreversibility analysis for Reheat-Regenerative power and

Rankine cycle. The irreversibility analysis was based on the concept of the second law of

analysis, and the results of the research revealed that incorporating feedwater heaters and

26

a reheat system would reduce the irreversibility by 12%. Aljundi (2009) performed a

study to determine the components having the largest exergy losses, and the study was

carried out for a steam power plant located in Jordan to identify the energy and exergy

losses. Hasan et al., (2009) analyzed a comparative study for nine different power plants

in Turkey; the research focused on the thermodynamic inefficiencies of each plant from

an exergetic and energetic viewpoint. A few researchers indentified that the major

exergetic loss is in the boiler and decided to focus only on the boiler (Rosen et al., 2008;

Bakhshesh and Vosough, 2012). There are also a number of studies carried out to identify

exergy loss for individual process components of power plants. Rosen (2001) studied a

comparative analysis of coal-fired power plants and nuclear steam power plants. The

study was focused on the energetic and exergetic losses, and it was found that the

exergetic efficiency is relatively higher in a coal-fired plant than in the nuclear steam

process.

Datta et al. (1999) analyzed the irreversibility in the boiler and found that

reducing the temperature gradient within the system can reduce irreversibility.

Regulagadda et al. (2010) performed a thermodynamic analysis of the boiler and turbine

operating at subcritical temperature and pressure. The results showed that the exergy loss

was found to be predominant in the boiler followed by the turbine. Li and Liu (2012)

performed an analysis based on the second law of thermodynamics for a 300 MW

thermal power plant. The analysis was based on the concept of the “fuel” and “product”

method. Li and Liu (2012) and Suresh et al. (2006) also reported that the boiler was

found to have the largest exergetic loss compared to the turbine and condenser.

27

Considering the facts stated above, performing a second law of thermodynamic

analysis (exergetic analysis) helps to resolve some specific inefficiencies of the power

plant. Thus, it helps to improve the overall efficiency of the power plant. It is very clear

from the above discussion that many researchers have worked on performing exergy

analysis for sub- and supercritical pulverized coal-fired power plants. However, no

research has been focused on exergy analysis of ultra supercritical pulverized coal-fired

power plants. Some studies have focused only on the furnace system operating at ultra

supercritical condition, and few methods were suggested to reduce the exergy losses

around the furnace. However, there are no statistical data available to prove that the

exergy loss could be reduced in the furnace system by implementing the methods

suggested. This is been the motivation of this thesis research, to proceed with further

development of ultra supercritical coal-fired power plant design and to perform the

exergy analysis in detail.

28

3. MODEL DEVELOPMENT

This chapter provides details of the process model developed in this study for the

simulation of coal-fired power plants. The model was built using the Microsoft® Visual

Basic Macro supported Excel® system to analyze the performance of a plant operating

under ultra supercritical conditions. The power plant model consists of a number of

process equipment modules that can be categorized into four groups including integrated

furnace system, series of turbines and pumps, condenser, and open and closed feedwater

heaters. The power plant model was built by incorporating the principles of coal

combustion, heat transfer, mass transfer, and thermodynamic properties. These principles

are discussed in detail in the following sections.

3.1 Integrated Furnace System

3.1.1 Flue Gas Composition

The combustion of coal generates heat energy for steam generation as well as

gaseous combustion products including CO2, H2O, O2, N2, SO2, and NO2. The amount of

oxygen gas (O2) required for complete combustion varies with elemental composition of

coal. According to the combustion reactions (Reactions 2.1 through 2.4), one mole of

carbon combines with one mole of oxygen gas to form CO2. One mole of hydrogen

combines with one-fourth mole of oxygen gas to form a half mole of H2O. One mole of

sulfur reacts with one mole of oxygen gas to form one mole of SO2. Similarly, one mole

of oxygen gas is needed to produce one mole of NO2. For a complete combustion of one

29

kilogram of coal, the total amount of O2 required is the sum of O2 consumed by individual

reactions. Thus, the total amount of O2 ( total, 2ON ) can be calculated from Equation 3.1.

i

OO N = N i, total, 22

(3.1)

where i, 2ON represents the amount of O2 required for combustion reaction i. In most

cases, there is a certain amount of elemental oxygen (O) in the fuel, which can be referred

to as fuel , ON . Therefore, the total amount of O2 supplied by air can be calculated using

Equation 3.2.

fuel, i, air , 2

22 O

i

OO NN = N (3.2)

Excess air generally refers to the amount of air supplied for combustion that is

more than the amount of air needed to fulfill the stoichiometric requirement. Excess air

helps to ensure that complete combustion is achieved (Bejan, 2006). Lack of excess air

will result in incomplete combustion, leading to the formation of toxic substances such as

carbon monoxide (CO). Based on the excess air percentage, the amount of flue gas

leaving the combustion varies. The calculation for excess air percentage depends on the

amount of O2 present in the flue gas and that in the air. The amount of excess oxygen

needed for complete combustion (

NO2 , excess) is a product of excess air percentage

( AirExcess % ) and amount of total O2 supplied by air.

).(100

) (%air , excess , 22 OO N

AirExcess = N

(3.3)

Therefore, the total amount of O2 actually supplied to the furnace can be written as

ai r , excess , suppl y , 222 OOO NN = N (3.4)

30

Based on the typical O2 content in the air (0.21 or 21 percent), the total amount of N2

entering the furnace can be expressed as

NN 2 , remaining = 10.21

0.21

NO 2 ,supply

(3.5)

Adding the amount of all gaseous components after combustion yields the total amount

of combustion products ( fluegasN ) as given in the following equation:

N fluegas = Ni

i

NN 2 , remaining NO 2 , supplyNCO2NH2O NSO2

NNO2 (3.6)

where Ni represents the number of moles of gaseous component i. Mole fraction of

component i ( iy ) can be determined by

fluegas

ii

N

Ny (3.7)

3.1.2 Combustion Temperature

To calculate the temperature of the combustion zone, it is necessary to determine

the specific heat of the flue gas. The formula to calculate the specific heat of the flue gas

is given below, where iCp, represents the molar specific heat of gaseous component i:

),( . )(, i

iif CpyCp (3.8)

The main source of energy for the furnace is from coal and preheated air. The heat

produced inside the furnace can be written as

..

).

.( AHcFurnace Q mLHVQ (3.9)

31

EVP Evaporator ECN Economizer SH Superheater RH Reheater

fw1 Feedwater in into economizer fw2 Feedwater in into evaporator

fw3 Steam in into Superheater T,f- Fuel temperature T3,f – T7,f Temperature of the flue gas

Figure 3.1 Scheme of steam flow inside furnace.

(Modified from Drbal et al., 1996 and Singer, 1991)

Coal Air

Flue gas in

Flue gas

out

Tf

T3,f

Combustion Zone

T5,f

T4,f

T7,f

T6,f

Heat supplied to EVP

Heat supplied to ECN

Heat supplied to RH

Heat supplied to SH

fw3

fw1

fw2

Superheated

steam

Reheated steam

To Reheat

Feedwater

ECN

EVP

RH

SH

32

The heat produced inside the furnace is translated into enthalpy of the combustion flue

gas, which can be calculated using the following equation:,

fT

Tf

ffluegasFurnace . dTCp. m.NQ C

,3

.

.

, (3.10)

where Cm.

represents the mass flow of coal, .

AHQ represents the amount of heat supplied

by the air preheater, and Tf and T3,f are feed temperature of coal and temperature of hot

flue gas or combustion zone, respectively. Since the molar specific heat values varies

with respect to the change in temperature, the values are assumed with in the range

(1.039-1.167) kJ/kmole.oC for model calulation (Smith et al., 1996). By assuming the

constant specific heat of flue gas, the temperature of the combustion zone can be written

as

ffluegas

Furnacef3,f

Cp. m. N

Q- T =

C ,

T

.

.

(3.11)

3.1.3 Air Preheater

The temperature of the preheated air from the preheater plays an important role in

determining the combustion zone temperature. The amount of heat supplied by the air

preheater can be calculated from

AHaaAH TCpm = Q ., ...

(3.12)

where

m.

a is the mass of air, Cp,a is the specific heat of air, and ∆TAH is the temperature

difference between inlet air and the preheated air.

33

3.1.4 Boiler Section

As can be seen from Figure 3.1, the boiler section of the integrated furnace

consists of a number of heat transfer modules designed for generating steam at different

pressures and temperatures. These heat transfer modules are economizer, evaporator,

superheater, and reheater. Energy balance equations associated with individual modules

are given below:

Economizer:

) ( . )( ., . 12

.

76

.

.

fwfww,f,fffluegasECN h- hmTTCpm. N = Q C (3.13)

Evaporator:

) ( . )( ., . 23

.

65

.

.

f wf ww, f, fff l uegasEVP h- hm =TTCpm.N = Q C (3.14)

Superheater:

) h- (hmTTCpm.N = Q s teamins teamouts,f,fffluegas SH C .)( ., ..

43

.

.

(3.15)

Reheater:

) h- (hmTTCpm.N = Q s teamins teamouts,f,fffluegasRH C .)( ., .

..

54

.

.

(3.16)

where sw mm.

,

.

denotes the mass rate of feedwater and steam. 2fw h , 1fw h denotes the

enthalpy of water leaving and entering the economizer, respectively. 3fw h denotes the

enthalpy of steam leaving the evaporator. steaminh , steamouth denotes the enthalpy of steam

in and out of the corresponding component . ,fT 3 , ,fT 4 , ,fT 5 , ,fT 6 , and ,fT 7 represent the

temperature of flue gas leaving the combustion chamber, leaving the superheater, leaving

the reheater, leaving the evaporator and leaving the economizer, respectively.

34



CND Condenser G Generator EVP Evaporator ECN Economizer SH Superheater RH Reheater

E1-E7 Expansion Valve

Figure 3.2 Scheme of pulverized coal-fired power plant with reference points.


FURNACE

Feed water

BP

E1 E2 E4 E5 E6 E7

E3

AH

Coal

Air

SH R

H

EV

P

ECN

DE

H1 H2 H3 H4 H5 H6 H7

HP IP LP

Saturated

Steam

Reheat Steam

CND

CP

Condensed Liquid

Furnace

G

37

36

21 27

26

23

20 17

12

14

13

16 19 22

24

2 6

9

7

F 32

33 E

A

B

C

D 34

35

28

29

30

1

31

5 8

11

15 18

4

9.1

12.1

3

10

25

16.1 19.1

22.1 6.1 2.1

35

Table 3.1 List of reference points used for the development of power plant.

Reference point Explanation

1 1st stage extraction from HP

2 Input to E1

3 Input to H1

4 Feedwater to ECN

5 2nd stage extraction from HP

6 Input to E2

7 Feedwater in to H2

8 1st stage extraction from IP

9 Feedwater input to E3

10 Input to H3

11 3rd stage extraction from IP

12 Input to E4

13 Feedwater input to H4

14 Feedwater input to DE

15 1st stage extraction from LP

16 Input to E5


18 2nd stage extraction from LP



21 3rd stage extraction from LP


23 Condensed steam in to H7

24 2nd stage extraction from IP

25 Input to BP from DE

26 Input to CP from CND

27 Last stage extraction from LP

28 Input to HP from SH

29 Input to IP from RH

30 Input to LP from IP

31 Input to RH from HP

32 Inlet air in to AH

33 Preheater air input to Boiler

34 Feedwater input to EVP

35 Feedwater input to SH

36 Cooling water in

37 Cooling water out

2.1 Steam return to H2


9.1 Steam return to DE




22.1 Steam return to CND

A Flue gas from combustion zone to SH

B Flue gas from SH to RH

C Flue gas from RH to EVP

D Flue gas from EVP to ECN

E Flue gas input to AH

F Stack gas out from AH

36

3.2 Turbines and Pumps

The total power generated from the power plant depends on the series of turbines,

such as high pressure turbines (HP), intermediate turbines (IP), or low pressure turbines

(LP). A typical arrangement of the turbines is illustrated in Figure 3.2. Note that small

portions of steam are extracted from these turbines to heat the feedwater heaters. This

results in the variation in flow rate of steam passing through each turbine section.

Therefore, the power produced from each turbine is derived from the combined power of

individual sections. The equation to calculate the total power from a turbine system can

be written as

(3.17)

where ,

.iHPW , ,

.iIPW , and ,

.iLPW represent the power output from a high pressure (HP),

intermediate pressure (IP), and low pressure turbine (LP), respectively. Considering the

turbine as an isentropic expansion process, the ideal enthalpy, idealout,h , at the turbine

outlet can be calculated for the corresponding entropy and pressure output for individual

turbine sections. The actual enthalpy, actu a lou t,h , for each turbine section can be written as

(Michael et al., 2011)

i ).( sentropicidealout,ininactualout, h h h= h (3.18)

where inh represents the enthalpy at the inlet of a turbine section and

isentropic is the

isentropic efficiency.

The net power output from the steam cycle is determined from the difference

between the total power produced from turbines and the sum of power required by the

both boiler feedwater pump and condenser pump. The net power then can be written as

o

i

LP

n

i

IP

m

i

iHPt ot alT WWW= W1

i,

1

i,

1

, ,

....

37

)W + W( - W =W BPCPtotalTNet

..

,

..

(3.19)

where

W.

BP and

W.

CP are the power required by the boiler feedwater pump and

condensate pump, respectively. Considering the condenser pump (CP) system operates

under steady flow, applying the first law of thermodynamics to calculate the amount of

work needed to drive the pump is given as

1,,

.

.)( PPW i nCPoutCPCP (3.20)

where outCPP , and inC PP , are the pressure at the exit and inlet of the condenser pump, and

1 denotes the specific volume of the condensate liquid. The power for the boiler

feedwater pump can be written as

2,,

.

.)( PPW i nBPoutBPBP (3.21)

where outBPP , , inBPP , are the pressure at exit and inlet of the boiler feed pump, and 2

denotes the specific volume of the boiler feed liquid.

Based on the new power ( NetW .

) value obtained, the net efficiency of the power plant can

be calculated on the HHV and LHV basis as follows:

HHV m

W =

c

Net

.

HHVNet

.

., (3.22)

LHV m

W =

c

Net

.

LHVNet

.

., (3.23)

The thermal efficiency of the steam cycle ( th ) can be determined from the amount of

heat supplied to the boiler and total net power generated from the steam cycle.

38

R H

.

S H

.

E V P

.

E CN

.

Net

.

B o iler

.

Net

.

th

Q Q Q Q

W

Q

W =

(3.24)

where Boiler

.

Q represents the heat input to the boiler system.

3.3 Condenser

The exit steam from the LP turbine is at relatively low pressure with low quality.

The condenser placed at the exit of the turbine receives this low quality steam. The

condenser is provided with external cooling water/air to support the transformation of

low pressure steam to condensed liquid. The energy balance around the condenser can be

written as

)( . ,,( . ,,

.

)

.

h hm=hhm i nwat eroutwat ercwi ncndoutcnds (3.25)

where cwm.

denotes the mass flow of cooling water and outc ndh , , incndh , are the enthalpy of

condensed steam at outlet and inlet. inw aterh , , outw aterh , represent the enthalpy of the cooling

water at the inlet and outlet, respectively.

3.4 Feedwater Heaters

The steam extracted from the turbine at different pressures heats the feedwater

inside the heater. For a known value of condenser pump discharge pressure, the

temperature and pressure of the water leaving the feedwater system can be calculated

using the following formula:

(3.26)

) ..

() ..

(1

1

out

n

i

iiin

m

i

ii h mh m

39

where .

im and ih represent the mass flow rate and enthalpy of stream (steam or liquid

water) i entering or leaving the heater. Table 3.2 provides a summary of the formulae

used in the power plant model.

3.5 Computational Algorithm for Power Plant Model

For any simulation of a power plant model, the algorithm is very important. The

algorithm used for simulating an ultra supercritical power plant is given in Figure 3.3.

The calculations begin by using the input parameters like coal composition, excess air

percentage, preheated air temperature, moisture content in coal,the main steam

temperature, the reheating temperature, mass flow of steam, steam extraction pressure of

turbines, and efficiency of the turbine and boiler. Once the input parameters are entered, a

calculation is carried out to determine the coal combustion products and their

compositions. Mass balance is applied to find the total moles of CO2, H2O, SO2, NOx, N2,

and O2 present after combustion. Specific heat values of combustion product and air are

determined to find the combustion zone temperature and furnace heat delivered to the

system. To proceed with the energy balance of the power plant, enthalpy values are

needed. Enthalpy is calculated as a function of temperature and pressure. In this study,

the steam property values were developed using Microsoft Excel® Macro, and the

equation used to calculate the enthalpy is given below:

Enthalpy calculation:

H=f (T, P) (3.27)

Pactual = P/Pn; Tactual =Tn/T (3.28)

40

where H, P, T are the enthalpy, pressure, and temperature. Pactual, Tactual represent the

actual pressure and actual temperature values. Pn, Tn are the co-efficients used to

determine the actual pressure and temperature based on the region n. For simplicity, four

different regions (n) were considered to categorize the co-efficient based on the

temperature and pressure range (Region 1 –Region 4 from (Appendix A)). The enthalpy

values were calculated for the temperature range 0oC to 2000

oC and pressure range

0.000611 MPa to 100 MPa. The equation given below was used to determine the

enthalpy with respect to the temperature and pressure.

Sum1 = n0 (k). J0 (k). Tactual J0 (k) – 1

(3.29)

Sum2 = n (k). Pactual i (k). j (k)

. Tactual j (k) – 1

(3.30)

where (J0, n0, i, j, n) co-efficients are selected based on the temperature and pressure

range from Appendix A. The final enthalpy equation is given as

H= 0.461526. T .Tactual. (Sum1+Sum2)

(3.31)

where i ranges from 1 to 4 (Appendix A).

Once the enthalpy is known, energy balance is performed on the components

inside the boiler section including the economizer, evaporator, superheater, and reheater.

Followed by the furnace calculation, outlet temperature of steam extraction from turbine,

workdone by the HP, IP, LP turbines The total power output from the turbine is

calculated from the below equation,

W.

total = W.

hp1 W.

hp2 W.

Ip1 W.

Ip 2 W.

Ip 3 W.

Lp1 W.

Lp2 W.

Lp3 W.

Lp4 (3.32)

The heat loss in the condenser is also calculated using the following equation,

Q.

CND m.

steamin . (hsteamout hsteamin) m.

waterin . (hwaterout hwaterin) (3.33)

41

To summarize, the energy balance, the relation for mass flow rate of coal input and net

efficiency of the power plant is given by the following equation,

Net, LHV = W

.

total W.

pump

m.

c . LHV (3.34)

The Thermal efficiency of the plant, and total workdone by the power plant were related

by the below equation.

th = W

.

total W.

pump

Q.

Boiler

(3.35)

42

Table 3.2 Enthalpy balance: formula used for power plant model.

Enthalpy balance formula

ECN

Q.

ECN m.

4 (h4 h34)Nf luegas .m.

C .Cp, f.(TE TF)

EVP

Q.

EVP m.


C .Cp, f.(TDTC)

SH

Q.

SH m.


C. Cp, f.(TB TA)

RH )T.(TC.m.N)h(hm.

Q BCp,f

.

fluegas 312931

.

RH C

HP ))h(h)m((m)h(hmW 51 1

.

28

.

128 HP 28

.

IP )mmm)(h(h)h(h)mm()h(hmW 24

.

8

.

29

.

1124248 8

.

29

.

829 29

.

IP ..

LP

W.

LP m.

30 (h30 h15) (m.

30 m.

15)(h 15 h18) (h18 h21).

(m.

30 m.

15 m.

18) (m.

30 m.

15 m.

18 m.

21) (h 21 h27)

H1

Q.

H1 m.

1h1m.

3h3 m.

4h4 m.

2 h2

H2

Q.

H2 m.

5 h5m.

7 h7 m.

2.1 h2.1m.

3h3m.

6 h6

H3

Q.

H3 m.

8 h8m.

9.1 h9.1 m.

6.1 h6.1m.

7h7m.

9 h9

H4

Q.

H4 m.

15 h15m.

13 h13 m.

14 h14 m.

12 h12

H5

Q.

H5 m.

15 h15m.

17 h17 m.

12. 1 h12. 1m.

13 h13m.

16 h16

H6

Q.

H6 m.

20 h20m.

18 h18 m.

16.1 h16.1m.

17 h17m.

19 h19

H7

Q.

H7 m.

21 h21m.

23 h23 m.

19.1 h19.1m.

20 h20m.

22 h22

CP 232623CP ) . P(PW

.

BP 102510BP ) . P(PW

.

CND

Q.

CND m.

27 . (h26 h27) m.

36 . (h37 h36)

DE

Q.

DE m.

24 h24 m.

14 h14 m.

9. 1 h9. 1m.

25 h25

43

Figure 3.3 Algorithm for ultra supercritical coal-fired power plant model with

reference to formula equation.

USC Power Plant model

Required Input Parameter Coal composition Excess air % Reheat temperature Efficiency of turbine, boiler Preheated air Main steam temperature Pressre of Boiler and condenser Turbine steam extraction pressure Pressure drop of boiler

Generate Thermodynamic Properties

Specific heat capacity for flue gas

Specific heat capacity of air

Enthalpy, entropy, saturation temperature,

Saturation pressure, vapor enthalpy

Estimate Coal Combustion Enthalpy Balance

Flue gas analysis Equation (3.6) Economizer Equation (3.13) CO2, H20, O2, NOx, SOx Evaporator Equation (3.14)

Energy produced by air preheater Equation (3.12) Superheater Equation (3.15)

Flue gas temperature Equation (3.11) Reheater Equation (3.16)

Steam Power Cycle

Outlet temperature of steam extraction from turbine Equation (3.18)

Workdone by HP, IP, LP turbines Equation (3.17) Workdone by Condenser pump, Boiler feed pump Equation (3.20) and (3.21)

Heat loss in condenser & Enthalpy around feedheater Equation (3.25) and (3.26)

Overall Performance of plant

Mass flow rate of coal Equation (3.23)

Net power output of plant Equation (2.12) Thermal efficiency of plant Equation (3.24)

Net efficiency of plant Equation (3.22)

44

3.6 Exergy Analysis for Individual Modules

A typical pulverized coal-fired power plant with specific reference points was

illustrated earlier in Figure 3.2. For any steady state process, the rate of exergy loss for

each component can be determined by calculating the exergy destruction rate that

depends on the entropy generated inside the system (Smith et al., 1996). Exergetic

efficiency will allow us to identify the irreversibility due to the heat, mass, and work

transfer. The formulas used to determine the exergy destroyed and exergetic efficiency

for the individual power plant components are discussed in the following sub sections.

3.6.1 Exergy Analysis of Integrated Furnace System

The exergy balance inside the furnace system can be written as

X ΨNmΨmΨmΨmΨmΨmΨm Fd

.

Ef l uegasc

.

2

.

.

28

.

31

.

4

.

ccaa ,..929

.

28314

..

(3.32)

where 28Ψ , 4Ψ , 92Ψ , 31Ψ , EΨ represents the exergy at superheater out, feedwater in,

reheater out, reheater in, and flue gas out from the economizer, respectively. The exergy

balance of the heat exchanger unit is given as

d, Boi l er

.

2

.

.

28

.

31

.

4

.

X ΨmΨmΨmΨm 929

.

28314 (3.33)

The exergetic efficiency of the furnace is given as

)()(

)()(

..

..

3131929428

, .

ΨNmΨmΨm

ΨmΨmΨmΨm

Efluegasc

.

ccaa

.

2

.

4

.

28

.

FurnaceΨ

(3.34)

45

3.6.2 Exergy Analysis of Turbines and pumps

The exergy balance for the turbines and pumps can be written as

.

X W- )Ψ .m)Ψ

.m TdtotalT

m

i

out

n

i

iisiniis ,,

.

1 1

,, . (. (

(3.35)

where i represents the steam extraction stage at any point of the turbine and iΨ denotes

the exergy of the stream i. The exergetic efficiency of the turbine is given as

)Ψ .m)Ψ

.m

W

m

i

o u t

n

i

iisiniis

to ta lT

.

Tu rb in eΨ

1 1

,,

,

. (. (

, . (3.36)

The exergy balance around the condenser pump can be written as

X W )Ψ(Ψm CPd

.

CP

.

OutCPinCP s

.

,,, (3.37)

where o u tCPinCP ΨΨ ,, , represents the exergy at the condensate pump inlet and outlet,

respectively. The exergetic efficiency of the condenser pump is given as

,,

,

CP

.

o u tCPinCP s

.

CPΨ

W

)Ψ(Ψm (3.38)

The exergy balance for the boiler feedwater pump can be written as

X W )Ψ(Ψm BPd

.

BP

.

outBPi nBP w

.

,,, (3.39)

where o u tBPinBP ΨΨ ,, , represents the exergy at the boiler feedwater pump inlet and outlet,

respectively.

The exergetic efficiency for the boiler feedwater pump can be given as

,,

, .

W

)Ψ(Ψm

B P

.

o u tB PinB P w

.

B PΨ

(3.40)

46

3.6.3 Exergy Analysis of Condenser

The exergy balance for the condenser can be written as,

d, CNDcw, Outcw, Incw s, Outs, Ins

.X)Ψ(Ψ

.m)Ψ(Ψ

.m (3.41)

where InsΨ , , outsΨ , , IncwΨ , , OutcwΨ , represents the exergy of steam in, steam out, cooling

water in, and cooling water out, respectively. The exergetic efficiency of the condenser is

given as

, .

)Ψ(Ψ.

m

)Ψ(Ψ.

m

s,Ou ts,Ins

cw,Ou tcw,incw

CNDΨ

(3.42)

3.6.4 Feedwater Heaters

The exergy balance for the feedwater heater can be written as

Hdouti

n

i

iini

m

i

i XΨmΨm ,

.

11

).().(

(3.43)

where

. im and i represents the mass flow rate and exergy of the stream (steam or

liquid water) i entering or leaving the heater. The exergetic efficiency of the feedwater

heater system is given as

ΨmΨm

ΨmΨm

o u two u tw

.

inwin w

.

o u tso u ts

.

insin s

.

H

,,,,

,,,,

(3.44)

where o u tso u ts

.

insin s

.

ΨmΨm ,,,,,, , , o u two u tw

.

inwin w

.

ΨmΨm ,,,,,, , denote the mass of steam in,

exergy of steam in, mass of steam out, exergy of steam out, mass of water in, exergy of

47

water in, mass of water out, and exergy of water out, respectively.

3.7 Overall Exergetic Efficiency of the Power Plant

The formula used to determine the exergetic performance for the power plant is

given below:

..

ΨmΨm

W

aacc

Net

.

Ψ,to ta l

(3.45)

where Ψ,total denotes the exergetic efficiency of the overall power plant, cΨ denotes the

exergy content in the coal used, and aΨ represents the exergy content in the air used.

3.8 Computational Model for Exergy Analysis

Simulation of exergy analysis is done after the development of a computational

model of the power plant. Figure 3.4 describes the computational flow of exergy analysis

developed for the power plant model shown in Figure 3.2. The input parameters required

to calculate the exergy of the power plant are reference temperature and reference

pressure. Entropy at each stage has to be known in order to proceed with the exergy

calculation. Equation 3.51 was used to determine the entropy of the individual component

in the power plant. The calculations performed to determine the entropy of the system are

given below:

Entropy calculation:

S=f (T, P)

(3.46)

48

where S denotes the entropy as a function of temperature T and pressure P. Pactual and

Tactual are calculated similarly to Equation 3.28. Equations 3.47 through 3.51 are used to

determine the entropy value.

Sum3 = n0 (k). J0 (k). Tactual J0(k) – 1

(3.47)

Sum4 = n (k). Pactual i(k) .j(k)

. (Tactual - 0.5) j(k) – 1

(3.48)

Sum5 = n0 (k). Tactual J0(k)

(3.49)

Sum6= n (k). Pactual i(k)

. (Tactual - 0.5) j(k)

(3.50)

S = 0.461526. (Tactual. (Sum5 + Sum6) - (Sum3 + Sum4)) (3.51)

where (J0, n0, i, j, n) co-efficients are selected based on the temperature and pressure

range from Appendix A.

Once the entropy values were calculated, exergy analysis was performed for the

integrated furnace system that includes an economizer, evaporator, superheater, and

reheater. After this, exergy analysis was performed on a series of turbines and pumps and

the condenser, air preheater, and feedwater heaters. The exergy destruction rate and

exergetic efficiency was determined for the individual components of the power plant

using the exergy values. The formulae used to determine the exergy destruction rate and

exergy efficiency are given in Table 3.3 and Table 3.4, respectively. The consolidated

results obtained after performing exergy analysis over the power plant yield the overall

exergetic efficiency of the plant and also help to identify the process equipment having

higher exergy loss rates.

49

Table 3.3 Exergy destruction: formula used for power plant model.

Exergy destruction formula

ECN ECNd

.X , )Ψ(Ψm 434

.

4

EVP EVPd

.X , )Ψ(Ψm 33 3

.

454

SH SHd

.X , )Ψ(Ψm 328 3

.

55

RH RHd

.X , )Ψ(Ψm 312931

.

HP HPd

.X , HP

.

311 1

.

28

.

12828

.

W) - Ψ(Ψ)mm()Ψ(Ψm

IP IPd

.X ,

IP

.

2 4

.

8

.

2 9

.

1 12 4

2 48 . 8

.

2 9

.

82 92 9

.

W)-mmm)(Ψ(Ψ

)Ψ(Ψ)mm()Ψ(Ψm

LP LPd

.X ,

LP

.

2721 21

.

18

.

15

.

30

.

18

.

15

.

30

.

2118181515

.

30

.

153030

.

W)-Ψ(Ψ)mmmm()mmm(

)Ψ(Ψ)Ψ)(Ψmm()Ψ(Ψm

H1 H1d

.X , )Ψ(Ψm)Ψ(Ψm 343

.

211

.

H2 2, Hd

.X ) Ψ(Ψm)Ψ(Ψm 737

.

655

.

H3 3, Hd

.X ) Ψ(Ψm)Ψ(Ψm 10710

.

988

.

H4 4, Hd

.X ) Ψ(Ψm)Ψ(Ψm 131413

.

121111

.

H5 5, Hd

.X ) Ψ(Ψm)Ψ(Ψm 171317

.

161515

.

H6 6, Hd

.X )(m)(m 2017 20

.

1918 18

.

H7 7, Hd

.X ) Ψ(Ψm)Ψ(Ψm 232023

.

222121

.

CP CPd

.X , CP2 62 32 3

. .W) Ψ(Ψm

BP BPd

.X , BP22

. .W) Ψ(Ψm 1 055

CND CNDd

.X , )Ψ .(Ψm)Ψ .(Ψm

.

272627

.

373636

DE DEd

.X , 2525

.

9. 19. 1

.

1414

.

2424

.

ΨmΨmΨmΨm

50

Table 3.4 Exergetic efficiency: formula used for power plant.

HP

)Ψ(Ψ)mm()Ψ(Ψm

W

311 1

.

28

.

128 28

.

HP

.

Ψ,HP

.

IP )Ψ(Ψ)mmm()Ψ(Ψ)mm()Ψ(Ψm

W

1124 24

.

8

.

29

.

248 8

.

29

.

829 29

.

IP

.

Ψ,IP

LP

)Ψ(Ψ)mmmm(

)Ψ)(Ψmmm()Ψ)(Ψmm()Ψ(Ψm

W

2721 21

.

18

.

15

.

30

.

211818

.

15

.

30

.

181515

.

30

.

153030

.

LP

.

Ψ,LP

Exergetic efficiency formula (%)

H1

ΨmΨm

ΨmΨm

22

.

11

.

33

.

44

.

Ψ,H

1 H2

ΨmΨmΨm

ΨmΨm

66

.

55

.

2.12.1

.

77

.

33

.

Ψ,H2

H3

ΨmΨmΨm

ΨmΨm

99

.

88

.

6.16.1

.

99

.

77

.

Ψ,H3

H4

ΨmΨm

ΨmΨm

1212

.

1111

.

1313

.

1414

.

Ψ,H4

H5

ΨmΨmΨm

ΨmΨm

1616

.

1515

.

12.112.1

.

1717

.

1313

.

Ψ,H5

H6

ΨmΨmΨm

ΨmΨm

1919

.

1818

.

16.116.1

.

2020

.

1717

.

Ψ,H6

H7

ΨmΨmΨm

ΨmΨm

2222

.

2121

.

19.119.1

.

2323

.

2020

.

Ψ,H7

DE

2424

.

9 .19 .1

.

2525

.

1414

.

Ψ,DE

ΨmΨm

ΨmΨm

ECN

Ψ(ΨNm

)Ψ(Ψm

EDfluegas c

.

.

Ψ,ECN

).

3444

EVP

Ψ(ΨNm

)Ψ(Ψm

DCflu eg a s c

.

.

Ψ ,EVP

).

3 53 43 4

SH

Ψ(ΨNm

)Ψ(Ψm

BAfluegas c

.

.

Ψ,SH

).

283535

RH

Ψ(ΨNm

)Ψ(Ψm

CBfluegas c

.

.

Ψ,RH

).

293131

CP

CP

.

262323

.

Ψ,CP

W

) Ψ(Ψm

BP 1 055

BP

.

2 2

.

Ψ,BP

W

) Ψ(Ψm

AH

)Ψ(ΨNm

) Ψ(Ψm

FEfluegasc

.

.

Ψ,AH

.

3332 32

CND

)Ψ(Ψm

) Ψ(Ψm.

.

Ψ,CND

2 62 7 .2 7

3 73 6 3 6

51

Figure 3.4 Algorithm for exergy analysis of power plant model with reference to formula

equation.

Start

Exergy Analysis

Required Input Parameter

Reference temperature

Reference pressure

Entropy of all process units

Exergy Result Analysis

Overall Exergetic efficiency of the plant Equation (3.45)

Spot the process equipment having higher exergy loss rate

Analyze factors responsible for the loss

Impose solution to reduce the exergy loss

Exergy Destruction of Furnace/Boiler Equation

(3.32) and (3.33)

Boiler

Heat exchanger

Economizer

Evaporator

Superheater

Reheater

Exergy Destruction of steam cycle (Table 3.3)

Turbine

Condenser

Condenser pump

Boiler feed pump

Feedwater heater

Exergetic Efficiency of Furnace/Boiler

(Table 3.4)

Boiler

Heat exchanger

Economizer

Evaporator

Superheater

Reheater

Exergetic Efficiency of steam cycle (Table 3.4)

Turbine

Condenser

Condenser pump

Boiler feed pump

Feedwater heater

52

3.9 Power Plant Model Validation

Today, there are a number of research studies focusing on the performance of

power plants, especially the efficiency of power generation. However, only a few studies

report the efficiency together with the complete collection of operating and design

conditions of the plant. These studies include works by Sanpasertparnich et al. in 2007

and by Wang et al. in 2012. Table 3.5 shows the input parameters required for simulating

a power plant model that includes coal consumption, equipment efficiency, pressure of

steam, and also temperature at different locations around the plant. These input values

were used in this study in order to simulate and validate the model developed here. Table

3.6 shows a comparison between the simulation outputs obtained from our model and

those reported in the literature. It is observed that the results obtained from the power

plant model developed in this study agree well with the results in the literature. This helps

to validate the power plant model developed for ultra supercritical conditions.

After the power plant model development, the exergy analysis was performed

using the second law of thermodynamics to study the performance of individual process

components. The exergetic performance of the power plant from the literature was

analyzed based on the reference points of a temperature of 25∘C and pressure of 0.103

MPa. The results of the exergy analysis were validated by comparing them with the

results obtained from Wang et al., 2012, and Aljundi, 2009. The power plant model

comparison was made using the same operating conditions as listed in Table 3.5 and

Table 3.7. Results were compared for the exergy destruction and exergetic efficiency of

the power plant. The simulation results for the exergy destruction, percent exergy

destruction of each component in the power plant is reported in Table 3.8 and Table 3.9.

53

Again, the exergy data obtained from the model developed in this study agreed well with

those from the literature, thereby validating the model.

54

Table 3.5 Power plant model validation – Input information from literature.

Description Wang et al., 2012 Sanpasertparnich et al.,

2007

Boiler efficiency (%)a 90 92

Turbine efficiency (%)a 90 92

Main steam temperature (∘C)

Main steam pressure (MPa)

571

25.4

600

25.3

Reheat temperature (∘C) 569 600

HP turbine

1st stage extract pressure (MPa)

2nd stage extract pressure (MPa)

IP turbine



3rd stage extract pressure (MPa)

LP turbine



3rd stage extract pressure (MPa)

4th stage extract pressure (MPa)

5th stage extract pressure (MPa)

6.79

4.3

2.05

1.04

-

0.437

0.133

0.065

0.020

0.005

4.31

3.29

2.52

1.27

0.9

0.22

0.07

0.03

-

-

Coal type

Carbon (wt%)

Hydrogen (wt%)

Oxygen (wt %)

Nitrogen (wt%)

Sulfur (wt%)

Moisture (wt%)

Ash (wt%)

Bituminous coal

-

57.52

3.11

2.78

0.99

2.00

2.10

23.70

Bituminous coal

(Illinois #6)

69

4.9

10

1

4.3

11.2

10.8

a values assigned in this study

55

Table 3.6 Power plant model validation case analysis.

Description Comparison between Wang et

al., 2012 and this study

Comparison between Sanpasertparnich

et al., 2007 and this study

Wang et

al., 2012

This

study

Absolute

%

deviation

Sanpasertparnich

et al., 2007

This

study

Absolute

%

deviation

Net power output

(MW)

671.0 698.0 4.02 350.0 347.0 0.86

Thermal

efficiency (%)

45.0 49.7 10.44 43.5 42.9 1.38

Net efficiency

(%)

42.0 46.2 10.00 32.3 32.0 0.93

Fuel consumption

(kg/s)

68.8 71.5 3.92 30.0 31.0 3.33

Flue gas

temperature (∘C)

1841 1898 3.10 1607 1650 2.68

56

Table 3.7 Exergy analysis validation –Literature operating parameters used.

Description Aljundi, 2009

Boiler efficiency (%)a 90

Turbine efficiency (%)a 90

Temperature of main steam (oC) 520

Pressure of main steam (MPa) 9.1

HP turbine


IP turbine


2nd

stage extract pressure (MPa)

LP turbine


2nd

stage extract pressure (MPa)

2.4

1.3

0.5

0.2

0.0

Boiler feed pump discharge pressure (MPa)

Condensate pump discharge pressure (MPa)

Pressure drop in feedwater heaters (%)a

Pressure drop in boiler (%)a

12.2

1.3

3.0

9.0

a values assigned in this study

57

Table 3.8 Exergy model validation case analysis - Exergy destruction comparison from

Wang et al., 2012.

Components

Exergy

destruction

(MW)

Reference

Exergy

destruction

(MW)

This model

Absolute

%

deviation

% Exergy

destruction

Reference

% Exergy

destruction

This model

Absolute

% deviation

Furnace (

Xd

., Furnace)

615.2 622.3 1.15 52.7 51.0 3.23

HP turbine 1st stage(

Xd

., HP1)

9.9 9.3 6.06 94.2 92.0 2.34

HP turbine 2nd stage(

Xd

., HP 2 )

3.0 3.0 0.00 93.2 93.0 0.21

IP turbine 1st stage (

Xd

., IP1)

3.6 3.9 8.33 96.3 96.0 0.31

IP turbine 2nd stage (

Xd

., IP 2)

2.6 2.6 0.00 96.4 96.2 0.21

LP turbine 1st stage (

Xd

., LP1)

2.7 1.8 33.33 96.2 96.5 0.31

LP turbine 2nd stage (

Xd

., LP 2)

2.1 2.8 33.33 97.2 97.0 0.21

LP turbine 3rd stage (

Xd

., LP 3)

2.3 2.9 26.09 93.3 93.1 0.21

LP turbine 4th stage (

Xd

., LP 4 )

11.4 13.6 19.30 75.8 75.7 0.13

LP turbine 5th stage (

Xd

., LP 5)

7.8 7.2 7.69 82.4 81.0 1.70

Feedwater heater1(

Xd

., FWH1)

1.1 1.2 9.09 95.4 95.4 0.00

Feedwater heater2(

Xd

., FWH 2)

1.5 1.7 13.33 94.0 94.0 0.00

Feedwater heater3(

Xd

., FWH 3)

2.2 2.4 9.09 89.4 89.2 0.22

Feedwater heater4(

Xd

., FWH 4 )

2.1 2.6 23.81 84.2 84.0 0.24

Feedwater heater5(

Xd

., FWH 5)

0.6 0.4 33.33 87.2 87.0 0.23

Feedwater heater6(

Xd

., FWH 6)

1.2 1.9 58.33 78.4 77.0 1.79

Feedwater heater7(

Xd

., FWH 7)

1.1 1.5 36.36 64.8 63.8 1.54

Condenser(

Xd

.,CND )

15.3 20.0 30.72 - 43.5 -

Boiler pump(

Xd

., BP )

1.7 1.7 0.00 89.4 87.0 2.68

Condenser pump(

Xd

.,CP)

0.1 0.1 0.00 71.6 71.0 0.84

Air preheater(

Xd

., AH )

16.9 20.0 18.34 86.7 86.2 0.58

Deaerator

Xd

., DE

2.0 2.5 25.00 80.5 80.0 0.62

Total 707.2 726.7 2.76

58

Table 3.9 Exergy model validation case analysis - Exergy destruction comparison from

Aljundi, 2009.

Components

Exergy

destruction

(MW)

Reference

Exergy

destruction

(MW)

This model

Absolute

%

deviation

% Exergy

efficiency

Reference

% Exergy

efficiency

This Model

Absolute

%

deviation

Furnace (

Xd

., F )

120.0 123.0 2.50 43.8 44.0 0.46

Turbine(

Xd

.,T )

20.0 20.4 2.00 73.5 72.6 1.22

Condenser(

Xd

.,CND )

13.7 14.0 2.19 26.4 26.7 1.14

Boiler pump(

Xd

., BP )

0.2 0.1 50 82.5 81.5 1.21

Condenser pump(

Xd

.,CP )

0.3 0.5 66.67 28.2 29.7 5.32

High pressure feedwater

heater1(

Xd

., FWH1)

0.4 0.5 25 97.4 97.6 0.21

High pressure feedwater

heater2(

Xd

., FWH 2)

0.3 0.3 0 95.3 96.9 1.68

Deaerator(

Xd

., DE )

0.3 0.1 66.67 95.3 95.7 0.42

Low pressure feedwater

heater1(

Xd

., LWH1)

0.3 0.3 0 89.5 89.0 0.56

Low pressure feedwater

heater2(

Xd

., LWH 2 )

0.2 0.3 50 67.3 66.1 1.78

Total 155.7 160.3 2.95

59

4. RESULTS AND DISCUSSIONS

4.1 Base Performance of an Ultra Supercritical Power Plant

This chapter provides insight into the performance of coal-fired power plants

operated under ultra supercritical conditions. Effects of operating parameters on plant

performance, especially exergy destruction and efficiency, are discussed so as to explore

the optimum operating conditions of the plants. Prior to the investigation of parametric

effects, base performance of the power plant was determined by simulating the developed

model using input parameters and their values listed in Table 4.1. Table 4.2 shows the

corresponding results obtained from the simulation. The base performance results include

net power output, rate of coal consumption, thermal efficiency of steam cycle, and net

efficiency of the power plant. Table 4.3 shows the exergy destruction and exergetic

efficiency of the base ultra supercritical plant. The exergy results are graphically

presented in Figure 4.1. Figure 4.2 shows the illustration of exergy loss flow from one

component to another.

The overall energy and exergetic efficiency of the power plant was found to be

41.12% and 36.19%, respectively. The exergy analysis revealed that the combustion

chamber has the least exergetic efficiency at 62.14%. The exergy analysis results

demonstrate that for the operating conditions considered in Table 4.1, 42.20% of exergy

could be lost by means of flue gas from the furnace without using an air preheater. This

shows that the furnace as the most significant exergy destruction component in the power

plant. The factors identified as the reasons for higher exergy loss are the fuel

composition, excess air percentage, moisture content in the fuel, and preheated air

60

temperature. The reason for the irrversibilty inside the combustion chamber is the

exchange of thermal energy between the reactants. The combustion process is said to be

ideal when there is no entropy generated during the process. The specific entropy of the

flue gas depends on the combustion temperature and pressure. Any effort to reduce the

average temperature difference between the hot and cold stream during the heat transfer

would result in the increase of exergetic efficiency. Adding a heat recovery unit, namely

an air preheater, next to the furnace helps to reduce the exergy loss from the flue gas. The

exergy loss reduced to 37.69% after the heat recovery process (see Figure 4.3).

From the distribution of the exergy destruction loss with respect to the input fuel

exergy in Figure 4.1, the exergy loss rate of the turbine is the next highest component

having greater exergy destruction. Maximum exergy loss in the turbine series occurs due

to the power generation. A high pressure turbine produces more power by consuming

17.52% of the exergy while an intermediate pressure turbine and a low pressure turbine

consume 9.91% and 21.42% of the exergy, respectively, to generate power. Destruction

in the feedwater heaters was found to be 1.30%. The exergy destruction rates in the

condensate pump and boiler feed pump were less predominant compared to total exergy

loss, representing 0.54% of the total loss. Although they were relatively small compared

with the total loss in the steam cycle, minimizing the exergy losses in these above

components will actually play an important part in the improvement of the exergy

efficiency of the ultra supercritical power plant. Exergy loss in the condenser is

significantly higher, where 7.04% of exergy was carried away by the cooling water.

Figure 4.2 shows the distribution of exergy rates inside the furnace. In summary, the best

61

gain in the power plant performance could be achieved by reducing the rate of entropy

generation in the combustion chamber.

62

Table 4.1 Input process parameters used for simulation of base power plant.

Parameters Value Unit

Temperature of main steam 700 oC

Pressure of main steam 35 MPa

Excess air 25 %

Feed temperature of cooling water 10 oC

Exit temperature of cooling water 30 oC

Preheated air temperature 300 oC

Flow rate of main steam 500 kg/s

Turbine discharge pressure

1st section of HP turbine

2nd

section of HP turbine

1st section of IP turbine

2nd

section of IP turbine

3rd

section of IP turbine

1st section of LP turbine

2nd

section of LP turbine

3rd


4th


4.5

3.5

3.0

1.2

0.9

0.43

0.25

0.12

0.008

MPa

MPa

MPa

MPa

MPa

MPa

MPa

MPa

MPa

Temperature of reheated steam 700 oC

Coal Composition

(Bituminous coal)

Carbon

Hydrogen

Nitrogen

Oxygen

Sulfur

Ash

Moisture

57.52

3.11

0.99

2.78

2.0

23.7

10

wt %

wt %

wt %

wt %

wt %

wt %

kg of H2O/ kg coal

63

Table 4.2 Base performance of ultra supercritical coal-fired power plant.

Performance parameters Value Unit

Net power generator output 828.59 MW

Rate of coal consumption 99.15 kg/s

Air in flow rate 746.24 kg/s

Flue gas temperature 1843.10 oC

Net energy efficiency (LHV based) 41.12 %

Thermal efficiency 53.98 %

Exergy efficiency 36.19 %

Flue gas composition (Mole basis)

CO2

N2

SO2

H2O

O2

NOx and others

Total

14.1

76.9

0.2

4.6

4.1

0.1

100

%

%

%

%

%

%

%

64

Table 4.3 Exergy destruction rate and exergy destruction percent of the plant.

Process

component

Exergy

in (MW)

Exergy

out

(MW)

Exergetic

destruction

(MW)

Exergetic

efficiency

(%)

% Total

destruction

HP turbine

IP turbine

LP turbine

Condenser

Pump

Combustion chamber

Boiler

Deaerator

Air preheater

High pressure feedwater heater

Low pressure feedwater heater

Total

354.59

241.39

386.40

171.02

132.63

2289.71

1165.28

59.84

316.68

180.46

106.55

346.45

234.10

322.72

150.95

124.61

1422.83

978.32

57.02

288.68

176.93

81.60

8.14

7.30

63.68

20.07

8.02

866.88

186.95

2.82

28.00

3.53

24.94

1220.35

97.70

96.98

83.52

88.26

93.95

62.14

83.96

95.29

91.16

98.04

76.59

0.67

0.60

5.22

1.65

0.66

71.04

15.32

0.23

2.29

0.29

2.04

100.00

65

8.14 7.3

63.68

20.07

8.02

866.88

186.95

28 24.94

1220.35

2.823.53

100

2.04

0.29

2.29

0.23

0.66

1.65

5.22

0.60.67

71.04

15.32

0.1

10

1000

HP tu

rbin

e

IP tu

rbin

e

LP turb

ine

Con

dens

er

Pump

Com

bustio

n ch

ambe

r

Boi

ler

Dea

erat

or

Air

preh

eate

r

Hig

h pr

essu

re h

eate

r

Low p

ress

ure he

ater

Total

cyc

le

Components

Exer

gy D

estr

uct

ion

(L

og s

cale

)

Exergy destruction rate (MW) Exergy destruction (%)

Figure 4.1 Distribution of exergy destruction for ultra supercritical power

plant.

66

2289.71

1422.83

708.95

369.69 338.72 305.8213.78

0

400

800

1200

1600

2000

2400

Inpu

t

Com

bustio

n pr

oduc

t

Super

heat

er

Reh

eate

r

Evapo

rato

r

Econo

miz

er

Air

preh

eate

r

Components

Ex

erg

y r

ate

(M

W)

Figure 4.2 Distribution of exergy rate inside furnace.

Base condition :

Preheated air temperature : 300oC Excess air : 25%

Mositure content in coal : 10%

67

Exergy loss due to flue gas out = 37.69% Exergy of coal in = 92.83%

Exergy loss due to condenser (cooling water) = 7.04% Exergy of air in = 0.44%

Exergy utilized to produce HP turbine power (HP) = 17.52% Power input to condensate pump (CP) = 0.28% Exergy utilized to produce IP turbine power (IP) = 9.91% Power input to boiler feed pump (BP) = 2.77%

Exergy utilized to produce LP turbine power (LP) = 21.42% Power input to cooling tower (CT) = 3.67%

Total exergy destroyed inside plant = 14.16%

Figure 4.3 Distribution of exergy percentage flow for ultra supercritical power plant. (Original in colour)


HPH High pressure feedwater heaters

BP Boiler feed pump


LPH Low pressure feedwater heaters

CP Condensate pump


DE Deaerator

AH Air preheater

CND Condenser

Ed (%) Exergy destructed

68

4.2 Parametric Effects on Exergy Loss

4.2.1 Effect of Moisture Content in the Coal

Table 4.3 represents the furnace as the main source for exergy destruction. One of

the factors responsible for this exergy loss in the furnace is the moisture content present

in the coal. The effect of moisture content can be observed from Figure 4.4 to Figure 4.6.

For example, Figure 4.4 shows the effect of moisture content in coal for different

preheated air temperature. The values were recorded at 15% excess air. The exergy

destruction was found to be 1022.55 MW at 10% moisture content and 350oC preheated

air, whereas the destruction rate was 1029.86 MW at 25% moisture content and 350oC

preheated air. These results indicate that there is a significant variation of nearly 7 MW of

exergy destruction rate for every 15% increase in the moisture content. This variation

could be decreased by increasing the preheated air temperature. For example, the exergy

destruction rate for 10% moisture content at 200oC and 350

oC was found to be 1069.97

MW and 1022.55 MW, respectively. This shows that the percentage of moisture content

increases the exergy destruction of furnace irrespective of the increase or decrease in the

preheated air temperature. As a result of the above discussion, it is clear that the moisture

content in coal has to be reduced to favour a more efficient power plant.

69

1020

1030

1040

1050

1060

1070

1080

1090

180 205 230 255 280 305 330 355

Exer

gy d

estr

uct

ion

(M

W)

10% Moisture content




Figure 4.4 Effect of moisture content at 5% excess air.

(Base conditions: Main steam temperature: 750oC, Main steam pressure: 35 MPa)

(Original in colour)

Preheated air temperature (oC)

70

1020

1030

1040

1050

1060

1070

1080

1090

180 205 230 255 280 305 330 355

Exer

gy d

estr

uct

ion

(M

W)









71

1010

1020

1030

1040

1050

1060

1070

1080

1090

180 205 230 255 280 305 330 355

Exer

gy d

estr

uct

ion

(M

W)









72

1020

1030

1040

1050

1060

1070

1080

1090

180 205 230 255 280 305 330 355

Exer

gy d

estr

uct

ion

(M

W)









73

1030

1040

1050

1060

1070

1080

1090

180 205 230 255 280 305 330 355

Exer

gy d

estr

uct

ion

(M

W)









74

4.2.2 Effect of Excess Air Percentage

From Figure 4.6 through Figure 4.8, for 350∘C preheated air temperature, it can

be observed that exergy destruction in the furnace is 1022.55 MW, 1029.26 MW, and

1038.88 MW for excess air at 15%, 20% and 25%, respectively. This illustrates the fact

that increase in excess air percentage increases the exergy destruction rate of the furnace.

The reason for this is that the total work performed by the combustion chamber reduces

as the flue gas temperature drops.

To determine the optimum amount of excess air required for combustion, the

percentage of excess air was varied from 5 – 25% to study its effect on the exergy

destruction rate of furnace. This analysis is illustrated in Figure 4.9 and Figure 4.10.

Figure 4.9 explains the variation of excess air percent and moisture content of coal with

the exergy destruction of the furnace at the minimum preheated air temperature (200oC).

The destruction rate decreased from 5 -16% at 10% moisture content, and further increase

in the excess air increased the rate of exergy destruction. The exergy destruction at 5%

excess air, 10% moisture content, and 200oC preheated air temperature was 1072.97 MW.

The exergy destruction at 16% excess air, 10% moisture content, and 200oC preheated air

temperature was 1069.97 MW. The exergy destruction at 25% excess air, 10% moisture

content, and 200oC preheated air temperature was 1076.90 MW. Therefore, the optimal

excess air percentage required for 200oC preheated air temperature and 10% moisture

content was 16%.

75

1065

1070

1075

1080

1085

1090

1095

1100

0 5 10 15 20 25 30

Exer

gy d

estr

uct

ion

(M

W)



Figure 4.9 Effect of excess air percent at 200oC preheated air temperature.



Minimum furnace

exergy destruction (MW)

Excess air percent

76

1010

1020

1030

1040

1050

1060

1070

1080

1090

0 5 10 15 20 25 30

Ex

erg

y d

estr

uct

ion

(M

W)



Figure 4.10 Effect of excess air percent at 350oC preheated air temperature.



Excess air percent

Minimum furnace

exergy destruction (MW)

77

4.2.3 Effect of Preheated Air Temperature

In this study the exergetic analysis of the furnace was completed by varying the

preheated air temperature from 200∘C to 350∘C. According to Woodruff et al. (2005), the

maximum temperature to which the air can be heated inside the preheater is 350∘C. This

analysis illustrates the effect of increasing the preheated air temperature with different

moisture content in coal. For example, Figure 4.7 shows that at 20% excess air and 10%

moisture content, the exergy destruction of the furnace is 1073.50 MW at 200∘C

preheated air temperature and 1029.26 MW at 350∘C preheated air temperature. This

demonstrates that heating the inlet air to a the maximum possible temperature decreases

the exergy destruction rate irrespective of the moisture content in the coal and the amount

of excess air.

78

4.2.4 Effect of Reheating Temperature

The exergy loss related to the turbine is significantly higher than that of the

furnace at low reheating temperature. The exergy distribution across the turbine is shown

in Table 4.3. It is evident that, among all the stages (HP, IP and LP), the low pressure

(LP) turbine has the lowest exergetic efficiency of 83.52% compared to the other turbine

stages. According to Habib et al. (1995), the maximum exergy efficiency of the turbine is

attained when the reheat pressure is maintained at 19% of the main steam pressure. In this

study, the parametric analayis was done by varying the reheat pressure from 6 – 20%.

The pressure range of the LP turbine inlet was varied from 6 – 35% for a better

understanding of the system’s performance. Figure 4.11 explains the effect of reheating

temperature on exergy efficiency in a turbine where the inlet pressure of the LP is

maintained at 0.7 MPa (35% of reheat pressure). The results were recorded for the

increase in reheating temperature from 500 to 750oC. Increase in reheat temperature

showed a positive impact on the exergy efficiency. For example, it is evident from Figure

4.11 that for a reheating temperature of 500oC and IP inlet pressure of 2.0 MPa (6

percent of main steam pressure), the exergy efficiency is 89.6%. Raising the reheat

temperature to 750oC led to the increase of efficiency of 91.7%. There was a significant

increase of 2% while the reheating temperature was increased as close as possible to the

main steam temperature. This shows that an increase in the reheating temperature

increases the exergy efficiency of the turbine. Therefore, it is essential to maintain the

highest possible reheating temperature to attain a more efficient power plant.

79

89

90

91

92

2 3 4 5 6 7 8

IP turbine inlet pressure (MPa)

Tu

rb

ine e

xerg

y e

ffic

ien

cy

(%

)

Figure 4.11 Effect of reheating temperature at 0.7 MPa LP inlet pressure.

(Base condition: Main steam temperature: 750oC, Main steam pressure: 35 MPa)


Reheat temperature (700∘C)


Maximum turbine exergy efficiency (%)





80

88

89

90

91

92

93

2 3 4 5 6 7 8


Tu

rb

ine e

xerg

y e

ffic

ien

cy

(%

)











81

88

89

90

91

92

93

94

2 3 4 5 6 7 8


Tu

rb

ine e

xerg

y e

ffic

ien

cy

(%

)











82

4.2.5 Effect of IP Turbine’s Inlet Pressure

The effect of inlet steam pressure at the intermediate pressure turbine for different

LP inlet pressures is shown in Figures 4.11 to 4.13. All three figures show that exergy

efficiency increases with an increase in the reheat or IP pressure (e.g., from 2.0 MPa to

3.4 MPa of IP inlet pressure in Figure 4.11, from 2.0 MPa to 3.8 MPa of IP inlet pressure

in Figure 4.12, and from 2.0 MPa to 4.0 MPa of IP inlet pressure in Figure 4.13 for

750∘C reheating temperature). The exergy efficiency is higher when the pressure ratio

between the inlet and outlet is higher. From Figure 4.13 it is shown that the turbine

reaches the maximum exergy efficiency of 93.4% at 0.9 MPa LP inlet pressure, 4.0 MPa

IP inlet pressure, and 750oC reheating temperature. Exploring the variation of IP turbine

inlet pressure from Figure 4.13 reaveals that increase or decrease in the LP turbine inlet

pressure beyond 4.0 MPa results in decrease of exergy efficiency. This shows that even

though an increase in the inlet pressure shows a positive result on exergy efficiency, there

is a decline in the exergy efficiency beyond a certain level. This level may be referred to

as the optimal level. This parameteric effect can also be observed in Figure 4.11 to Figure

4.13.

83

4.2.6 Effect of LP Turbine Inlet Pressure

From Figures 4.11 to 4.13, we can establish the optimal relationship between the

inlet pressure of the IP turbine and the LP turbine. From Figure 4.13, it is shown that the

turbine exergy efficiency is 93.2% at 0.9 MPa LP inlet pressure, 2.0 MPa IP inlet

pressure, and 750oC reheating temperature, and from Figure 4.12, the exergy efficiency is

92.6% at 0.8 MPa LP inlet pressure, 2.0 MPa IP inlet pressure, and 750oC reheating

temperature. Exploring the variation of LP inlet pressure with respect to IP inlet pressure

from Figure 4.14 reveals that an increase or decrease in the LP inlet pressure beyond 0.9

MPa results in a decrease in turbine exergy efficiency. Performing an analysis camparing

the exergy efficiency with the pressure range helps to conclude the optimum pressure

required for the process.

4.2.7 Effect of LP Turbine Exit Pressure

The effect of decreasing the outlet pressure of the LP turbine shows much less

variation in terms of exergy destruction and exergetic efficiency. Increase in the pressure

ratio between the inlet and outlet stream results in higher exergetic efficiency. From

Figure 4.15, at 2.0 MPa IP inlet pressure, 0.9 MPa LP inlet pressure, and 0.006 MPa LP

exit pressure, the exergetic efficiency is 93.4%. Increasing the LP exit pressure to 0.007

MPa results in decrease of turbine exergy efficiency at a very minimal level of 93.14%.

Therefore, increasing the LP exit pressure by 0.001 MPa results in the decrease of

exergetic efficiency by 0.05 percent. The exergy destruction at this stage also depends on

the quality of the exit steam.

84

86

87

88

89

90

91

92

93

94

2 3 4 5 6 7 8 9

Intermediate turbine inlet pressure (MPa)

Tu

rbin

e ex

erg

y e

ffic

ien

cy (

%)

Figure 4.14 Effect of IP inlet pressure with varying LP inlet pressure.



RHT = 750oC, LP = 0.9 MPa

RHT = 750oC, LP = 0.8MPa













85

91

92

93

94

0.005 0.009 0.013 0.017

LP turbine exit pressure (MPa)

Tu

rb

ine e

xerg

y e

ffic

ien

cy

(%

)

Figure 4.15 Effect of LP exit pressure with varying LP inlet pressure.

(Base condition: Main steam temperature: 750oC, Main steam pressure: 35 MPa, IP inlet

pressure: 2.0 MPa)

LP inlet pressure = 0.7 MPa

LP inlet pressure = 0.8MPa

LP inlet pressure = 0.9 MPa

86

4.3 Optimum Operating Conditions

Implementing ideal ultra supercritical power plant steam parameters depends on

the boiler design conditions. In this study, it was assumed that the boiler was made up of

nickel-based super alloys, which meets the requirement of holding the steam temperature

at 750oC and pressure at 35 MPa (Viswanathan et al., 2003). The percentage of excess air

fed into the furnace was assumed in the range from 5 to 25 percent. The amount of

acceptable moisture content in the coal was optimized for the imput range from 10 - 25

percent. Optimizing the inlet and outlet pressure of the IP and LP turbines requires the

assumption of a pressure range to determine the highest possible exergy efficiency. The

assumptions for pressure range are illustrated in Table 4.4. The main steam flow rate to

the high pressure turbine is maintained at a constant 500 kg/s. The boiler efficiency and

isentropic efficiency of the turbines were also maintained at a constant 92% to study the

performance of the system.

Table 4.5 gives the consolidated results for the optimum conditions that could be

used to obtain a more exergy efficient power plant. Incorporating the above parameters

into the simulation yields a net power output of 941.25 MW. The net exergetic efficiency

of the power plant was found to be 41.18%. Comparing the optimized result with the base

results given in Table 4.2 shows an increase of 112.65 MW of net power and 4.9% of

exergy efficiency.

87

The effect of exergy destruction on a furnace with excess air percentage was

found to be minimal at 15% for 350oC preheated air temperature and 10% moisture

content. Any further increase or decrease in the excess air percentage increased the

exergy destruction rate and decreased the exergetic efficiency of the furnace. Increasing

the preheated air temperature to the maximum attainable temperature proved best for

achieving an efficient result. The lower the moisture content in the coal, the better the

system’s exergetic performance. Reduction in the exergy destruction of the turbine series

could be achieved by increasing the IP inlet pressure to the optimum level for a constant

LP inlet pressure. In this study, it was observed that 12% of the main steam pressure (4.0

MPa) is sufficient for the reheating pressure to attain a maximum exergy efficiency of

93.4%. From Figure 4.14, it is clear that maintaining the reheating temperature as close as

possible to the main steam temperature results in a better performance of the turbine. The

optimum conditions for the ultra supercritical power plant are described in Figure 4.16.

88

Table 4.4 Optimal process parameters for ultra supercritical power plant.

Components Range

considered

Value Unit

Main steam temperature 700-750 750 oC

Main steam pressure 30-35 35 MPa

Excess air 5-25 15 %

Turbine discharge pressure

HP 1st stage

HP 2nd

stage

IP 1st satge

IP 2nd

stage

IP 3rd

stage

LP 1st stage

LP 2nd

stage

LP 3rd

satge

LP 4th

stage

3-5

5-7

1-3

0.5-1

0.5-1

0.2-1

0.05-0.2

0.05-0.2

0.005-0.05

4

5

2.0

1.0

0.9

0.43

0.25

0.12

0.006

MPa

MPa

MPa

MPa

MPa

MPa

MPa

MPa

MPa

Reheat temperature 700-750 750 oC

Reheat pressure 2-7 4 MPa

Preheated air temperature 250-350 350 oC

Boiler pump pressure 30-35 30.8 MPa

Condensate pressure 1.5-1.8 1.8 MPa

Moisture content in coal 10-25 10 %

89

Table 4.5 Optimal results obtained for ultra supercritical power plant.

Overall plant’s results Value Unit

Net power output 941.25 MW

Net efficiency 44.16 %

Flow rate of coal 96.24 kg/s

Thermal efficiency 55.23 %

Overall Exergy efficiency 41.18 %

Exergy results Exergy

destruction (MW)

Exergy

efficiency

%

Combustion chamber 611.55 72.45

Boiler 275.20 79.21

High pressure turbine 8.31 97.83

Intermediate pressure turbine 7.30 96.98

Low pressure turbine 51.73 86.19

High pressure feedwaterheater 5.87 96.97

Low pressure feedweater heater 24.89 76.59

Pump 8.02 93.95

Condenser 20.07 88.26

Deaerator 2.82 95.29

Air preheater 14.51 95.08

90

Exergy loss due to flue gas out = 34.04% Exergy of coal in = 92.68%

Exergy loss due to condenser (cooling water) = 6.94% Exergy of air in = 0.42% Exergy utilized to produce HP turbine power (HP) = 18.32% Power input to condensate pump (CP) = 0.28%

Exergy utilized to produce IP turbine power (IP) = 9.31% Power input to boiler feed pump (BP) = 2.85%

Exergy utilized to produce LP turbine power (LP) = 12.83% Power input to cooling tower (CT) = 3.77% Total exergy destroyed inside plant = 18.55%

Figure 4.16 Distribution of exergy percentage flow for optimized ultra supercritical power plant. (Original in colour)


HPH High pressure feedwater heaters

BP Boiler feed pump


LPH Low pressure feedwater heaters

CP Condensate pump


DE Deaerator

AH Air preheater

CND Condenser

Ed (%) Exergy destructed

91

5. CONCLUSIONS AND FUTURE WORK

A second law of thermodynamic analysis was performed for a ultra

supercritical power plant to explore the performance of individual components. From the

exergy analysis results obtained, the furnace was identified as having the highest exergy

destruction. The turbines had the second highest exergy destruction rate. Some of the

conclusions regarding the exergy destruction of different components and options to

reduce the exergy destruction in the furnace and turbine are given below:

Exergy destruction rate was greater in the combustor than the heat exchanger unit.

The exergy loss in the furnace was 886.75 MW. The factors responsible for the

furnace exergy losses were excess air percentage, preheated air temperature, and

moisture content in the coal.

The exergetic efficiency increased with decrease in the moisture content of coal

and increased in accordance with preheated air temperature. Increase in the excess

air beyond 15% showed an increase in the exergy destruction rate.

The exergy efficiency of the turbine is 93.4%. The exergy loss depends purely on

the pressure range of the input and outlet stages of the turbine series. The

exergetic efficiency increased in accordance with an increase in the reheating

temperature.

Increase in the IP turbine inlet pressure showed an increase in the exergetic

efficiency until the pressure reached the optimum value. Beyond the optimum

value, the efficiency decreased with further increase in pressure. The optimum

92

pressure at the IP turbine inlet was 4.0 MPa for 0.9 MPa IP outlet pressure. The

exergy loss increased with decrease in the LP inlet pressure.

The following are the recommendations for the future work:

For the power plant to operate at a very high temperature and pressure, it is

important to focus more on the plant design. An advanced material is required to

design a steam turbine that can withstand temperatures above 700∘C and

pressures above 30 MPa.

The steam turbine has to be designed in such a way that it matches the furnace

conditions. Future research should also focus on identifying a coating that can

withstand the high temperature and pressure and protect the equipment from

steam oxidation and erosion inside the turbines. Also, the material properties

throughout the overall design of an ultra supercritical power plant using coal as

the fuel should be investigated.

93

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100

APPENDIX

APPENDIX A

Table A.1 Co-efficiencts for region 1 to calculate the enthalpy and entropy

(Wagner and Kretzschmar, 2008)

Temperature >= 273.15 and Temperature <= 623.15

Pressure > Psat and Pressure <= 100

Region 1: Pn =16.53, Tn=1386

Co-efficients used for region1

J0 = (0, 1, -5, -4, -3, -2, -1, 2, 3)

n0 = (-9.69, 10.08, -0.0056, 0.07, -0.40, 1.42, -4.38, -0.28, 0.021)

i = (0, 0, 0, 0, 0, 0, 0, 0, 1, 1, 1, 1, 1, 1, 2, 2, 2, 2, 2, 3, 3, 3, 4, 4, 4, 5, 8, 8, 21, 23,

29, 30, 31, 32)

j = (-2, -1, 0, 1, 2, 3, 4, 5, -9, -7, -1, 0, 1, 3, -3, 0, 1, 3, 17, -4, 0, 6, -5, -2, 10, -8, -

11, -6, -29, -31, -38, -39, -40, -41)

n = (0.14, -0.84, -3.75, 3.38, -0.95, 0.15, -0.01, 8.12E-04, 2.83E-04, -6.07E-04, -

0.018, -0.032, -0.021, -5.28E-05, -4.71E-04, -3.00E-04, 4.76E-05, -4.41E-06, -

7.26E-16, -3.16E-05, -2.82E-06, -8.52E-10, -2.28E-06, -6.51E-07, -1.43E-13, -

4.05E-07, -1.27E-09, -1.74E-10, -6.87E-19, 1.44E-20, 2.63E-23, -1.191E-23,

1.82E-24, -9.35E-26)



101




Pressure > 0 and Pressure <= Psat

Or

Temperature > 623.15 and Temperature <= 863.15

Pressure > 0 and Pressure <= Psat

Or

Temperature > 863.15 and Temperature <= 1073.15

Pressure > 0 and Pressure <= 100

Region 2: Pn =1, Tn=540


J0 = (0, 1, -5, -4, -3, -2, -1, 2, 3)

n0 = (-9.69, 10.08, -0.0056, 0.07, -0.40, 1.42, -4.38, -0.28, 0.021)

i = (1, 1, 1, 1, 1, 2, 2, 2, 2, 2, 3, 3, 3, 3, 3, 4, 4, 4, 5, 6, 6, 6, 7, 7, 7, 8, 8, 9, 10, 10,

10, 16, 16, 16, 18, 20, 20, 20, 21, 22, 23, 24, 24, 24)

j = (0, 1, 2, 3, 6, 1, 2, 4, 7, 36, 0, 1, 3, 6, 35, 1, 2, 3, 7, 3, 16, 35, 0, 11, 25, 8, 36,

13, 4, 10, 14, 29, 50, 57, 20, 35, 48, 21, 53, 39, 26, 40, 58)

n = (-1.77E-03, -0.017, -0.04, -0.05, -0.05, -3.30E-05, -1.89E-04, -3.93E-03, -

0.04, -2.66E-05, 2.04E-08, 4.38E-07, -3.22E-05, -1.50E-03, -0.04, -7.88E-10,

1.27E-08, 4.8E-07, 2.29E-06, -1.67E-11, -2.11E-03, -23.89, -5.90E-18, -1.26E-

06, -0.03, 1.12E-11, -8.23, 1.98E-08, 1.04E-19, -1.02E-13, -1.00E-09, -8.08E-11,

0.10, -0.33, 8.91E-25, 3.06E-13, -4.20E-06, -5.90E-26, 3.78E-06, -1.271E-15,

7.30E-29, 5.54E-17, -9.43E-07)



102



Temperature >= 623.15 and Temperature <= Tsat

Pressure >= Psat and Pressure <= 100

Region 3: Pn =333, Tn=647.096


J0 = (0, 1, -5, -4, -3, -2, -1, 2, 3)

n0 = (-9.69, 10.08, -0.0056, 0.07, -0.40, 1.42, -4.38, -0.28, 0.021)

i = (0, 0, 0, 0, 0, 0, 0, 0, 1, 1, 1, 1, 2, 2, 2, 2, 2, 2, 3, 3, 3, 3, 3, 4, 4, 4, 4, 5, 5, 5, 6,

6, 6, 7, 8, 9, 9, 10, 10, 11)

j = (0, 0, 1, 2, 7, 10, 12, 23, 2, 6, 15, 17, 0, 2, 6, 7, 22, 26, 0, 2, 4, 16, 26, 0, 2, 4,

26, 1, 3, 26, 0, 2, 26, 2, 26, 0, 1, 26)

n = (1.06, -15.73, -20.94, -7.68, 2.61, -2.80, 1.20, -8.45E-03, -1.26, -1.15, -0.88, -

0.64, 0.38, -85,4.89, -3.05, 0.03, 0.12, -0.27, 1.38, -2.01, -8.21E-03, -0.47, 0.04, -

0.44, 0.90, 0.70, 0.10, -0.32, -0.50, -0.02, 0.094, 0.16, -0.01, -0.014, 5.79E-04, -

3.23E-03, 8.09E-05, -1.65E-04, -4.49E-05)



103




Pressure >= 0 and Pressure <= 100

Region 4: Pn =1, Tn=1000


J0 = (0, 1, -3, -2, -1, 2)

n0 = (-13.17, 6.85, -0.02, 0.36, -3.11, -0.32)

i = (1, 1, 1, 2, 3)

j = (0, 1, 3, 9, 3)

n = (-1.25E-04, 2.17E-03, -0.004, -3.97E-06, 1.29E-07)



exergy analysis of pulverized coal-fired ultra

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