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 4.5 Effect of moisture content at 10% excess air 70
Figure 4.6 Effect of moisture content at 15% excess air 71
Figure 4.7 Effect of moisture content at 20% excess air 72
Figure 4.8 Effect of moisture content at 25% excess air 73
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 4.12 Effect of reheating temperature at 0.8 MPa LPT inlet pressure 80
Figure 4.13 Effect of reheating temperature at 0.9 MPa LPT inlet pressure 81
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
HP High pressure turbine H1-H3 High pressure closed feedwater heaters BP Boiler feed pump IP Intermediate pressure turbine H4-H7 Low pressure closed feedwater heaters CP Condensate pump
LP Low pressure turbine DE Deaerator AH Air preheater
CND Condenser G Generator
Figure 1.3 Scheme of supercritical pulverized coal-fired power plant.
(Modified from Bakhshesh and Vosough, 2012)
10
HP High pressure turbine H1-H4 High pressure closed feedwater heaters BP Boiler feed pump IP Intermediate pressure turbine H5-H8 Low pressure closed feedwater heaters CP Condensate pump
LP Low pressure turbine DE Deaerator AH Air preheater
C Condenser G Generator
Figure 1.4 Scheme of ultra supercritical pulverized coal-fired power plant.
(Modified from Bakhshesh and Vosough, 2012)
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
HP High pressure turbine H1-H3 High pressure closed feedwater heaters BP Boiler feed pump IP Intermediate pressure turbine H4-H7 Low pressure closed feedwater heaters CP Condensate pump
LP Low pressure turbine DE Deaerator AH Air preheater
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.
(Modified from Bakhshesh and Vosough, 2012)
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
17 Feedwater input to H5
18 2nd stage extraction from LP
19 Feedwater input to E6
20 Feedwater input to H6
21 3rd stage extraction from LP
22 Feedwater input to E7
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
6.1 Steam return to H3
9.1 Steam return to DE
12.1 Steam return to H5
16.1 Steam return to H6
19.1 Steam return to H7
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.
34 (h34 h35)Nf luegas .m.
C .Cp, f.(TDTC)
SH
Q.
SH m.
35 (h28 h35)Nf luegas .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
1st stage extract pressure (MPa)
2nd stage extract pressure (MPa)
3rd stage extract pressure (MPa)
LP turbine
1st stage extract pressure (MPa)
2nd stage extract pressure (MPa)
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
1st stage extract pressure (MPa)
IP turbine
1st stage extract pressure (MPa)
2nd
stage extract pressure (MPa)
LP turbine
1st stage extract pressure (MPa)
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
section of LP turbine
4th
section of LP turbine
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)
HP High pressure turbine
HPH High pressure feedwater heaters
BP Boiler feed pump
IP Intermediate pressure turbine
LPH Low pressure feedwater heaters
CP Condensate pump
LP Low pressure turbine
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
15% Moisture content
20% Moisture content
25% 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)
10% Moisture content
15% Moisture content
20% Moisture content
25% Moisture content
Figure 4.5 Effect of moisture content at 10% excess air.
(Base conditions: Main steam temperature: 750oC, Main steam pressure: 35 MPa)
(Original in colour)
Preheated air temperature (oC)
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)
10% Moisture content
15% Moisture content
20% Moisture content
25% Moisture content
Figure 4.6 Effect of moisture content at 15% excess air.
(Base conditions: Main steam temperature: 750oC, Main steam pressure: 35 MPa)
(Original in colour)
Preheated air temperature (oC)
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)
10% Moisture content
15% Moisture content
20% Moisture content
25% Moisture content
Figure 4.7 Effect of moisture content at 20% excess air.
(Base conditions: Main steam temperature: 750oC, Main steam pressure: 35 MPa)
(Original in colour)
Preheated air temperature (oC)
73
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
15% Moisture content
20% Moisture content
25% Moisture content
Figure 4.8 Effect of moisture content at 25% excess air.
(Base conditions: Main steam temperature: 750oC, Main steam pressure: 35 MPa)
(Original in colour)
Preheated air temperature (oC)
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)
10% Moisture content
25% Moisture content
Figure 4.9 Effect of excess air percent at 200oC preheated air temperature.
(Base conditions: Main steam temperature: 750oC, Main steam pressure: 35 MPa)
(Original in colour)
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)
10% Moisture content
25% Moisture content
Figure 4.10 Effect of excess air percent at 350oC preheated air temperature.
(Base conditions: Main steam temperature: 750oC, Main steam pressure: 35 MPa)
(Original in colour)
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)
(Original in colour)
Reheat temperature (700∘C)
Reheat temperature (750∘C)
Maximum turbine exergy efficiency (%)
Reheat temperature (650∘C)
Reheat temperature (600∘C)
Reheat temperature (550∘C)
Reheat temperature (500∘C)
80
88
89
90
91
92
93
2 3 4 5 6 7 8
IP turbine inlet pressure (MPa)
Tu
rb
ine e
xerg
y e
ffic
ien
cy
(%
)
Figure 4.12 Effect of reheating temperature at 0.8 MPa LP inlet pressure.
(Base condition: Main steam temperature: 750oC, Main steam pressure: 35 MPa)
(Original in colour)
Maximum turbine exergy efficiency (%)
Reheat temperature (750∘C)
Reheat temperature (700∘C)
Reheat temperature (650∘C)
Reheat temperature (600∘C)
Reheat temperature (550∘C)
Reheat temperature (500∘C)
81
88
89
90
91
92
93
94
2 3 4 5 6 7 8
IP turbine inlet pressure (MPa)
Tu
rb
ine e
xerg
y e
ffic
ien
cy
(%
)
Figure 4.13 Effect of reheating temperature at 0.9 MPa LP inlet pressure.
(Base condition: Main steam temperature: 750oC, Main steam pressure: 35 MPa)
(Original in colour)
Reheat temperature (750∘C)
Reheat temperature (700∘C)
Reheat temperature (650∘C)
Reheat temperature (600∘C)
Reheat temperature (550∘C)
Reheat temperature (500∘C)
Maximum turbine exergy efficiency (%)
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.
(Base condition: Main steam temperature: 750oC, Main steam pressure: 35 MPa)
(Original in colour)
RHT = 750oC, LP = 0.9 MPa
RHT = 750oC, LP = 0.8MPa
RHT = 750oC, LP = 1.0 MPa
RHT = 750oC, LP = 0.7 MPa
RHT = 750oC, LP = 1.4 MPa
RHT = 750oC, LP = 1.6 MPa
RHT = 750oC, LP = 2.0MPa
RHT = 500oC, LP = 0.7 MPa
RHT = 500oC, LP = 0.8MPa
RHT = 500oC, LP = 0.9 MPa
RHT = 500oC, LP = 1.0 MPa
RHT = 500oC, LP = 1.4 MPa
RHT = 500oC, LP = 1.6 MPa
RHT = 500oC, LP = 2.0 MPa
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)
HP High pressure turbine
HPH High pressure feedwater heaters
BP Boiler feed pump
IP Intermediate pressure turbine
LPH Low pressure feedwater heaters
CP Condensate pump
LP Low pressure turbine
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
REFERENCES
Aljundi, I.H. (2009). Energy and exergy analysis of a steam power plant in jordan.
Applied Thermal Engineering, 29, 324- 328.
Bakhshesh, M. and Vosough, A. (2012). Boiler parametric study to decrease
irreversibility. Indian Journal of Science and Technology, 5(4), 2534-2539.
Bejan, A. (2006). Advanced Engineering Thermodynamics, 3rd Edition. John Wiley &
Sons: New York, NY,USA.
Chaibakhsh, A. and Ghaffari, A. (2008). Simulation modeling practise and theory.
Simulation Modelling Practice and Theory, 16, 1145–1162.
Casarosa, C; Donatini, F. and Franco, A. (2004). Thermo economic optimization of heat
recovery steam generators operating parameters for combined plants. Energy, 29,
389-414.
Central Electricity Authority. CEA(2011a), Highlights of power sector in India:
http://www.cea.nic.in/reports/monthly/executive_rep/mar11/1-2.pdf.
(16 April 2011).
Central Electricity Authority. CEA(2011b), Power supply position in India 2010-11:
http://www.cea.nic.in/reports/monthly/executive_rep/mar11/24-25.pdf.
(16 April 2011).
Central Electricity Authority. CEA(2012), Power supply position in India 2010-11:
http://www.electricity.ca/media/Industry%20Data%20and%20Electricity%20101
%20May%202012/ElectricityGenerationinCanadabyFuelType1990_2011.pdf.
(24 April 2012).
94
Cheng, L.; Guo, J.I.F. and Mingtian, X.u. (2010). Thermodynamic analysis of waste heat
power generation system. Institute of Thermal Science and Technology, Energy,
35,2824-2835.
Datta, A. and Som, S.K. (1999). Energy and exergy balance in a gas turbine combustor.
J Power Energy—Proc Inst Mech Eng, 213A, 23–32.
Drbal, L.F; Boston, P.G.; Westra, K.L. and Erickson, R.B. (1996). Power Plant
Engineering. Ist Edition, Black & Veatech, Springer, New York.
Energy Information Administration, EIA. (2011). International Energy outlook:
http://www.eia.gov/countries/country-data.cfm?fips=IN&trk=p1,
(13 March 2011).
Energy Information Administration, EIA. (2012). International Energy outlook:
http://www.eia.gov/electricity/annual/html/epa_a_03.html, (6 August 2012).
Energy Information Administration, EIA. (2013). International Energy outlook:
http://www.eia.gov/totalenergy/data/monthly/index.cfm#coal, (30 January 2013).
Gwosdz, A.; Leisse, A. and Quenders, H.J. (2005). Pulverized coal firing system for the
operation of steam generators with low excessive air. VGB Powertech. 85(11),
67-73.
Habib, M.A. and Zubair, S.M. (1992), 2nd-law-based thermodynamic analysis of
regenerative - reheat rankine-cycle power plants. Pergamon-Elsevier Science
Ltd. Energy, 17, 295-301.
Habib, M.A.; Said, S.A.M. and Al-Zaharna, I. (1995), Optimization of reheat pressuers in
thermal power plants. Pergamon-Elsevier Science Ltd. Energy, 20(6), 555-565.
95
Hasan, H.E.; Ali, V.A; Burhanettin, A.D.; Suleyman, H.S.; Bahri, S.; Ismail, T.; Cengiz,
G. and Selcuk, A. (2009). Comparative energetic and exergetic performance
analyses for coal-fired thermal power plants in turkey. International Journal of
Thermal Sciences, 48, 2179–86.
International Atomic Energy Agency, IAEA. (2004). Vienna International Centre:
http://www.iaea.org/Publications/Magazines/Bulletin/Bull461/power_to_the_peo
ple.html, (10 september 2012).
International Energy Agency, IEA. (2012). Power generation from coal:
http://www.iea.org/aboutus/faqs/coal/, (8 september 2012).
Jayamaha. L. (2008). Energy efficient building systems. Handbook: 1st Edition. Mcgraw
Hill eduaction, Europe.
Kakaras, E.; Ahladas, P. and syrmopoulos, S. (2002). Computer simulation studies for the
integration of dryer to coal power plant. Fuel, 81(5), 583-593.
Kaushik, S.C.; SivaReddy, V. and Tyagi, S.K. (2011). Energy and exergy analyses of
thermal power plants. A review, Renewable and Sustainable energy reviews, 15,
1857-1872.
Kjaer, S. (2002). The advanced supercritical 700oC pulverized coal-fired power plant.
VGB Power Tech, 7, 47-49.
Kiga, T.; Yoshikawa, K.; Sakai, M. And Mochida, S. (2000). Characteristics of
pulverized coal combustion in high temperature preheated air. Journal of
propulsion and power, 16(4), 601-605.
Kitto, J.B. (1996). Developments in pulverized coal-fired boiler technology, Babcock &
Wilcox.
96
Lako, P. (2004). Coal-fired power technologies: Coal-fired power options on the brink of
climate policies (ECN-C-04-076), ECN Policy studies, Netherlands.
Leung, P. and Moore, R.E. (1966). Closure to discussion of cycles for supplying steam to
desalting evaporators of dual-purpose power-generation plants. ASME J. Eng.
Power, 88, 283.
Li, Y. and Liu, L. (2012). Exergy analysis of 300 MW coal-fired power plant.
International Conference on future Electrical Power and energy systems, Energy
Procedia, 17, 926- 932.
Lior, N. and Zhang, N. (2007). Energy, exergy and second law performance criteria.
Energy, 32, 3.
Marion, J.L.; Liljedahl, G.N. and Black, S. (2004). A Review of the state of the art and a
view of the future for combustion based coal power generation. The 29th
International technical Conference on Coal Utilization & Fuel Science, April
18-22, Clearwater, Florida.
Meyer, L.; Tsatsaronis, G.; Buchgeister, J. and Schebek, L. (2009). Exergo environmental
analysis for evaluation of the environmental impact of energy conversion
systems. Energy, 34, 75-89.
Michael, J.M.; Howard, N.S.; Daisie, D.B. and Margaret, B.B. (2011). Fundamentals of
engineering thermodynamics Handbook: 7th Edition. John Wiley & Sons, Inc.
Nag, P.K. and De, S. (1997). Design and operation of a heat recovery steam generator
with minimum irreversibility. Applied Thermal Engineering, 17, 8.
Oktay, Z. (2009). Investigation of coal-fired power plants in turkey and a case study:
Can plant. Applied Thermal Engineering, 29, 550–557.
97
Perry, R.H.; Green, D.W. and Maloney, J.O. (1997), Perry’s Chemical Engineer’
Handbook: 7th Edition. McGraw-Hill, New York.
Raja, A.K.; Amit, P.S. and Manish, D. (2006). Power Plant Engineering Handbook.: Ist
Edition. New Age International Limited, 1-10.
Reddy, V.; Ramkiran, G.; Kumar, K.A. and Nag, P.K. (2002). Second law analysis of a
waste heat recovery steam generator. International Journal of Heat and Mass
Transfer, 45, 1807-1814.
Regulagadda, P.; Naterer, G.F. and Dincer, I. (2010). Exergy analysis of a thermal power
plant with measured boiler and turbine losses. Applied Thermal Engineering, 30,
970–976.
Rosen, M.A. (2001). Energy and exergy-based comparison of coal-fired and nuclear
steam power plants. International Journal of Exergy, 1, 180–192.
Rosen, M.A.; Dincer, I. and Kanoglu, M. (2008). Role of exergy in increasing efficiency
and sustainability and reducing environmental impact. Energy Policy, 36,128–
137.
Sanpasertparnich, T. (2007). Monte carlo simulation of pulverized coal-fired power
plants: Efficiency improvement and CO2 capture options. International journal
of greenhouse gas control, 1, 143-150.
Sengupta, S.; Datta, A. and Duttagupta, S. (2007). Exergy analysis of coal based 210
MW thermal power plant. International Journal of Energy Research, 31,14-28.
Singer, J.G. (1991). Combustion fossil power: a reference book on fuel burning and
steam generation: 4th generation. combustion engineering power systems group,
Windsor, Connecticut.
98
Smith, J.M.; VanNess, H.C. and Abbott, M.M. (1996). Introduction to chemical
engineering thermodynamics: 5th Edition. McGraw-hill, New York.
Sue, D.C. and Chuang, C.C. (2004). Engineering design and exergy analyses for
combustion gas turbine based power generation system. Energy, 29, 1183–1205.
Suresh, M.V.J.J.; Reddy, K.S. and Ajit, K.K. (2006). Energy and exergy analysis of
thermal power plants based on advanced steam parameters. National conference
on advances in energy research, India: IITB.
Thess, A. (2011). The entropy principle-Thermodynamics for the unsatisfied Handbook:
2nd Edition. Institute of thermodynamics and Fluid mechanics, 43-45.
Tsatsaronis, G. and Cziesla, F. (2011). Energetic and exergetic analysis of complex
systems, Energy and exergy system analysis and optimization- Volume I.
Encyclopedia of life support systems (EOLSS).
U.S. Department of energy, U.S.DOE. (1999). Market Based Advanced Coal Power
Systems. Washington, DC:
http://energy.gov/search/site/1999, (10 April 2011).
U.S. Department of energy, U.S.DOE. (2005). Market Based Advanced Coal Power
Systems. Washington, DC:
http://energy.gov/search/site/2005, (10 April 2011).
Viswanathan, R.; Romanosky, R.; Purgert, R. and Johnson, H. (2003). Boiler materials
for USC plant. Electric power research institute:
http://www.netl.doe.gov/publications/proceedings/03/materials/Viswanathan.pdf,
(1 January 2013).
99
Wang, L.; Yang, Y.; Morosuk, T. and Tsatsaronis, G. (2012). Advanced thermodynamics
analysis and evaluation of a supercritical power plant. Energies 2012, 5, 1850-
1863.
Wei, D.H.; Lu, X.S.; Lu, Z. and Gu, J.M. (2007) . Performance analysis and optimization
of organic rankine cycle (ORC) for waste heat recovery. Energy Conversion and
Management , 48(4),1113-1119.
World Resources Institute (WRI). (2012). Publications-Climate, Energy and trnasport,
Washington.
http://insights.wri.org/topic/oil-and-gas, (10 January 2012).
Wagner, W. and Kretzschmar, H.J. (2008). International steam tables based on the
industrial formulation 1AP WS-1F97: 2nd edition, Springer.
World Resources Institute (WRI). (2007). Publications-Climate, Energy and trnasport,
Washington:
http://www.wri.org/publication/wri-annual-report-2006, (13 December, 2011).
Woodruff, E.B.; Lammers, H.B. and Lammers, T.F. (2005). Steam Plant Operation
Handbook: 8th Edition. McGraw-Hill, New York.
Yunus, A.C. and Michael, A.B. (2008). Thermodynamics Handbook : 6th Edition.
McGraw-Hill, New York.
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
Table A.2 Co-efficiencts for region 2 to calculate the enthalpy and entropy
(Wagner and Kretzschmar, 2008)
Temperature >= 273.15 and Temperature <= 623.15
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
Co-efficients used for region2
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
Table A.3 Co-efficiencts for region 3 to calculate the enthalpy and entropy
(Wagner and Kretzschmar, 2008)
Temperature >= 623.15 and Temperature <= Tsat
Pressure >= Psat and Pressure <= 100
Region 3: Pn =333, Tn=647.096
Co-efficients used for region3
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
Table A.4 Co-efficiencts for region 4 to calculate the enthalpy and entropy
(Wagner and Kretzschmar, 2008)
Temperature >= 1073.15 and Temperature <= 2273.15
Pressure >= 0 and Pressure <= 100
Region 4: Pn =1, Tn=1000
Co-efficients used for region4
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)