DESIGN OF
GAS TURBINE
COMBINED CYCLE AND
COGENERATION SYSTEMS
May 2009
Maher A. Elmasri
Thermoflow Inc.
______________________________________
©Copyright 1990-2009
Notes are copyrighted and may not be reproduced
in whole or in part without written permission.
Preface
This book’s content is the same as the “Design of Gas Turbine Combined Cycles and Cogeneration Systems”
seminar, which I taught about seventy times in twenty years, 1990 through 2009. About three thousand
engineers from around the world attended these 3-day sessions and received versions of this book. Many
additional printed copies were sold or given to those requesting them, but the book was never turned over to
a mass-media publisher and has remained in limited circulation of a few thousand printed copies.
Friends have asked why this was never turned over to a mass-media publisher. There are several reasons.
First, the content was evolving with the result that the entire book was always like Mr. Micawber’s goal of an
entire suit of new clothes at once, in the Dickens novel “David Copperfield”. By the time one chapter was
updated, another had gone out-of-date. This was due to my primary focus on the development and
expansion of the Thermoflow software suite, leaving no time to update the entire book at once. Second, the
instability and boom-bust cycles in the thermal power generation industry, and the economy as a whole, had
a disruptive effect. Whenever a boom period produced high seminar attendance at frequent seminars,
resulting in renewed enthusiasm to update much of the book and widely publish it, a bust ensued and
thwarted the momentum. Third, is aversion to dealing with the corporate mass publishing industry.
Since the last time I taught the seminar in 2009, many have asked why it was no longer being offered. The
answer is partly in the preceding paragraph, but there are additional reasons. When I started the seminar in
1990, combined cycles were not mainstream, and many experienced plant designers had much better
understanding of coal and nuclear Rankine Cycle plants (which then generated 70% of US electricity) than
combined cycle plants (which then generated less than 10% of US electricity). When I last taught the seminar
in 2009, combined cycles had become mainstream, their principles more widely understood, and were on
their way to generating 30% of US electricity. Also, when I started the seminar in 1990, GT PRO® was a barebones calculation tool, DOS-based, no user-friendly graphics, and with limited built-in expertise and features.
By 2009, thanks to my brilliant colleagues at Thermoflow, especially Dr. Gwo-Tung Chen and Dr. Patrick
Griffin, GT PRO® had become a highly user-friendly combination of calculation tool and expert knowledge
repository. Over time, the knowledge taught in the seminar and explained in this book was embedded into
the Thermoflow software suite. Practically-useful exergy analysis of combined cycles, pioneered in my 1980’s
papers, and covered to the extent time permitted in the seminar and its book was integrated into GT PRO®, so
any user had full access to this technique and knowledge. My models for the cooled gas turbine, published in
the 1980’s, were embedded into the THERMOFLEX® program as a “Cooled Turbine Stage” component
available to all users of this program. Many design and modeling procedures for steam turbines, condensers,
and cooling towers, described in the seminar, had been integrated into the Thermoflow software suite by
2009. The methodology of selecting cogeneration cycle configurations was embedded into PDE®. The logic of
parsing through and analyzing different repowering configurations was built into its own dedicated
Thermoflow program, RE-MASTER. Thus, in effect, the Thermoflow software suite, with its greatly increased
user-friendliness, and its inclusion of much of the content of the seminar and this book, has subsumed the
seminar and this book as a repository of plant design knowledge, analysis, and modeling.
Another reason for my waning enthusiasm to repeat the seminar, is the changing audience. At the earlier
seminars, the audience included many highly experienced engineers. Those new to combined cycles had a
solid grasp of Rankine Cycle practice, and understood how to analyze and calculate thermal power cycles.
This enabled them to quickly understand the content that was new to them. Some were gas turbine experts
at major OEMs who certainly understood their gas turbines better than I, but came to learn about combined
cycle optimization, or steam turbines, or HRSGs. Some were HRSG, or steam turbine, or cooling tower experts
from major OEMs, who certainly knew their specialty and its details far better than I, but came to learn about
the complete cycle and its other components. With this audience profile, it was likely that for much of the
content, there was someone in the audience who knew more about the current topic than I, and many of
those contributed valuable comments from which the rest of the audience and I learnt many new things.
However, by the latter seminars, the profile had increasingly tilted towards an inexperienced and
undereducated audience, with many struggling to follow or understand the content or its basic concepts; and
this made teaching it more challenging and less stimulating.
When I started the seminar in 1990, I invited my friend and colleague, the late R. W. (Dick) Foster-Pegg, to
teach portions of it. Dick was a great engineer, with an immense grasp of engineering principles and
applications, despite minimal theoretical training. He complemented my (at that time) academically-biased
background. Dick was 32 years my senior, had worked at Rolls-Royce on airplane piston engines during World
War II, on gas turbines starting in the 1950’s, and on practical engineering of combined cycles in the 1960’s, at
Bechtel then at Westinghouse, until retiring in the 1980s. I had earned my Ph.D. on gas turbine cooling at MIT
in the late 1970’s, then as a professor at MIT in the 1980’s did extensive original work, supervised many
theses, and wrote numerous publications on practically applying fundamental thermodynamics and heat
transfer to gas turbine and combined cycle analysis and optimization. The diversity between Dick’s
background and generation and mine, created a unique learning experience for those who attended the
seminar until Dick retired from participating in 2000; and from 2000 through 2009, I taught it alone.
The copyright notice at the beginning of each chapter indicates the vintage of its content. Chapters 1-3, 5-7,
and 9-18 are my original content from 1990 with some updates and revisions over the years (year of latest
revision as shown). Chapters 4, 8, 19 and 20 are my newer content written in 1999 and 2000 on subjects
which had been taught by Dick in earlier seminars. Chapters 21 and 22 were written in 2008 on CO2 capture
and solar thermal energy, topics that had not been considered earlier.
Chapters 1 and 2, which focus on basic applied thermodynamics and cycle analysis, include original
fundamental insights and interpretations that had not been previously published. They draw heavily on the
work I had done at MIT, much of it published in papers during the 1980’s and referenced at the end of
Chapter 2. Some who attended the seminar reiterated these insights and analyses in subsequent
publications, with a few unfortunately omitting reference to the seminar and its book they received.
Although this book has not been updated at all since 2009, and some chapters since 2002, its fundamental
insights and approach remain as relevant as ever. The main differences between current practice and the
book’s content will be found in the chapters dealing with gas turbines. The principles are as valid as always,
but higher firing temperatures, better materials, and improved aerodynamics have added a few percentage
points to both gas turbine and combined cycle efficiencies.
Maher A. Elmasri
Jacksonville, FL
March, 2019
 M. A. Elmasri, 1990-2008
Chapter 1: Introduction
INTRODUCTION
Revised August, 2008
© Maher Elmasri 1990-2008
1.1
HEAT & WORK
Heat is disorganised energy, manifested in matter by increasing the intensity of the random motion of its
constituents. The kinetic theory of gases describes thermal energy in a gas by the intensity of motion of
its molecules, oscillating and vibrating randomly in all directions. Thermal energy is manifested in a
solid by the intensity of vibration of its crystals within their lattices. Thus, heat is a random form of
energy. It is diffuse, clumsy, and difficult to harness as organised motion.
Work is organised energy; force that can push uniformly in one direction. Because it is organised, it can
lift or move things, and is highly convenient to use and transmit.
Heat, the easy, simple, poor form of energy, has been in widespread use by people since prehistoric times.
Since recorded history began, wood or other combustible fuel would be burnt to generate heat for
cooking, drying or melting ore to extract metals. People still consume large amounts of heat, to keep a
house warm, or cook, or dry clothes, etc; as well as in most industrial processes.
Work, the well-organised form of energy, was not available to people in large amounts until the heat
engine came into widespread use. Until about 200 years ago, the sources of work were human power,
animal power, wind-mills, and the few water-mills built on dammed rivers. The basic transportation
device was the sailing ship, which could harness the natural force of wind, as it occurs, to derive useful
work from it.
A physically-fit man can generate work for periods of several hours at rates of 30 - 60 watts. A worldclass athlete can generate as much as 400 watts for periods of minutes. Thus if an average, fit, person
worked fairly hard on a physical job, 8 hours per day, 250 days per year, he could perform approximately
100 kWh of physical work all year ($8 worth of electricity !).
To estimate human consumption of work in pre-heat-engine days, we note that a horse or other draft
animal may have performed work at the rate of about 1/3rd of a kW for 2000 hours per year, producing
about 600 kWh of work per year. Assuming one draft animal per ten people of population, one fit hardworking person per three of population, and a little wind or water power, we find that the amount of work
used per capita per year was about 100 kWh.
Widespread deployment of the practical heat engine, beginning in the first quarter of the 19th century,
revolutionised life by enabling people to use amounts of work greatly in excess of these historic levels. In
the USA in 2007, the per capita consumption of energy in the form of work was over 18,000 kWh,
roughly 180 times the historic standard which prevailed from ancient times up to the 1800's. The recent
improvement, over the last two centuries, in mankind's physical quality of life has been derived from the
basic fact that people's ability to lift and move objects has increased 180-fold. The danger to our natural
environment, created by our increased ability to rearrange it, can also be attributed to this basic fact.
1.1.1
ESTIMATED GLOBAL HEAT & WORK CONSUMPTION (2007)
Worldwide total consumption of energy as heat by the end-user in 2007 is estimated at roughly 260
Quads (Quadrillion BTU), i.e. 260 x 1015 BTU/year. This estimate is extrapolated from the 2005 data in
Reference [1], after subtracting from primary energy the amounts of fossil fuels used for electricity
generation and transportation, as well as the amount of nuclear heat used for electricity, and the amount of
1-1
 M. A. Elmasri, 1990-2008
Chapter 1: Introduction
hydro electricity generated from the total global primary energy sources. With a worldwide population of
6.6 billion people in 2007, this comes to an average of 11,400 kWh(thermal)/capita.
Worldwide electricity generation in 2007 is estimated at 16.7 x 1012 kWh. With a world population of 6.6
billion, and assuming 6.67% T&D losses, one gets a per capita worldwide electricity consumption of
about 2,800 kWh/year per person. Worldwide use of fossil fuel energy for transportation in 2007 is
estimated at 98 x 1015 BTU/year. If one were to assume an overall efficiency of 20% for transportation
machinery on the average, one would conclude that worldwide use of transportation Work was roughly
870 kWh/capita, averaged over the world's population. Thus the estimated worldwide consumption of
Work per capita is roughly 3,670 kWh, or about 37 times that of the pre-heat-engine era.
Fig. 1 summarises estimated global Heat & Work Consumption for 2007.
Total 341 Quads
Total 15,070 kWh/capita
Figure 1. Estimated 2007 global end-use energy consumption, based on data of Reference [1] and
approximations by the author (Quad = 1015 BTU – Global population = 6.63 billion)
Naturally, the large multiple of work consumed by comparison with historic levels is not all “work” in the
strict sense, i.e. it is not all motive (force x distance), since it includes all uses of electricity, such as
lighting, some heating, cooling, etc. However, the fact remains that even these thermal and non-motive
uses of electricity enhance the quality of life as much as motive work, since the availability of clean,
organised energy from a wire allows it to be used in myriad, convenient ways, that would be very difficult
to achieve from a direct source of heat. Just think of how inconvenient it would be if you had to light a
fire every time you needed to read in the dark or toast a piece of bread, or be unable to refrigerate food.
1.1.2
ESTIMATED HEAT & WORK CONSUMPTION IN THE USA (2007)
US consumption of energy as heat in 2007 is estimated at roughly 29 Quads (Quadrillion BTU), i.e. 29 x
1015 BTU/year. This estimate is derived from the data in Reference [2], by subtracting the amounts of
fossil fuels used for electricity generation and transportation from the total fossil fuel consumption.
Please note that this data is based on Higher Heating Value (HHV) of fossil fuels. With a population of
305 million people in 2007, this comes to an average of 28,200 kWh(thermal)/capita.
In 2007, the electricity consumed by the end-users in the USA was 3.89 x 1012 kWh [2], which, with a
2007 population of 305 million, gives about 12,700 kWh per year per-capita. Work used by the average
person, in the form of electricity alone, is one hundred and thirty times that estimated for the pre-heatengine era. In addition to Work consumed as electricity, we consume a vast amount of Work propelling
cars, trucks, airplanes, ships, etc. Assuming 20% overall efficiency of converting fuel heat (HHV) to
work in the transportation sector, we find that the US, which consumed approximately 29 x 1015
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 M. A. Elmasri, 1990-2008
Chapter 1: Introduction
BTU/year worth of transportation fuel in 2007, expends about 5,600 kWh of Work per capita per year for
transportation. Thus the total US consumption of Work per capita is about 18,300 kWh/year, 180 times
the estimated level for people in the pre-heat-engine era.
Fig. 2 summarises US Heat & Work Consumption in 2007.
Total 48 Quads
Total 46,500 kWh/capita
Figure 2. Estimated 2007 US end-use energy consumption, based on data of Reference [2] and
approximations by the author (Quad = 1015 BTU – US population = 305 million)
1.1.3
ENERGY COSTS IN THE USA (2007)
Based on data in Reference [2], in 2007 the 29 Quads consumed in the US in the form of heat had an
average end-user price of about $13/MMBTU, or 4½ cents/kWh(th) across the residential, commercial
and industrial sectors. The cost was about $bn 386/year, or about $1,265 per capita per year.
The US 2007 retail cost of electricity (to the end user), averaged about 10.6 cents/kWh for residential use
(37%), 9.7 cents/kWh for commercial use (36%), and 6.4 cents/kWh for industrial use (27%). Averaged
over residential, commercial & industrial use, it comes to about 9.1 cents/kWh. The total end-user value
of electricity consumed is about $bn 356/year, or about $1,170 per capita per year.
The average 2007 “on-highway price” of motor fuels, according to Reference [2] was in the range $2.80$3.03 per gallon. For the purposes of estimating the value of Work used in transportation, we shall
assume a 2007 price of $3 per US gallon. A gallon provides 125,000 BTU of heat (HHV), which
corresponds to a price of $24/million BTU, or 8.2 c/kWh of thermal energy. Thus, with the assumption
of 20% transportation efficiency, the cost of Work used in transportation is about 41 cents/kWh, and the
cost of transportation energy per capita of about $2,280 per year.
Figure 3 summarises the above data.
In summary, about 39.5% of USA total energy consumption is in the form of Work; its cost, however, is
73% of the total energy cost. Work on average, cost its consumer about 19 c/kWh, a little over four times
the average cost of 4.5 c/kWh for heat.
Work in the form of electricity cost the average end-user about 9.1 c/kWh, including the infrastructure
capital costs of generation, transmission and distribution.
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 M. A. Elmasri, 1990-2008
Chapter 1: Introduction
Total $4,700/capita
Total $bn 1,440
Figure 3. Estimated 2007 US end-use energy costs, based on Ref. [2] and approximations by the author
Work for transportation, cost the average end-user an estimated 41 c/kWh for fuel alone, in addition to the
capital cost of the vehicles (cars, trucks, airplanes, trains, ships), paid by the end-user and the
infrastructure cost (roads, airports, etc..), paid partly by the end-user but largely by society.
1.1.4
ELECTRICITY GENERATION SOURCES
The figure below shows that on a global basis, as of 2005, two-thirds of the electricity generated was from
conventional fossil fuels: coal (40%), gas (20%), and petroleum (7%). Renewable energy accounted for
about 18%, mostly hydro. Nuclear energy was the source for the remaining 15%.
For the USA, as of 2007, 72% of the electricity generated was from conventional fossil fuels: coal (49%),
gas (21%), and petroleum (2%). Renewable energy accounted for about 9%, mostly hydro. Nuclear
energy accounted for the remaining 19%.
Hydro
16.4%
Other
1.8%
Coal
40.3%
Nuclear
15.2%
Petroleum
6.6%
Natural Gas
19.7%
Fig. 4. Global electricity generation by energy source (2005)
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 M. A. Elmasri, 1990-2008
Chapter 1: Introduction
Hydro
6.0%
Other
2.6%
Nuclear
19.4%
Coal
48.6%
Petroleum
1.9%
Natural Gas
21.5%
Fig. 5. US electricity generation by energy source (2007)
Primary fossil fuel energy consumed to generate electricity in the USA (fuel heat input, HHV, of coal,
petroleum and natural gas) was 29.4 x 1015 BTU/year [2],and the electricity generated therefrom was 2.98
x 1012 kWh. This implies an average generation efficiency of about 36.2% (HHV), corresponding to
about 39% LHV when one assumes average fuel chemical properties.
The costs of all fossil fuels used to generate electricity in the USA in 2007 was about $bn 89. For an
average kWh of electricity generated from fossil fuels, approximately 3 cents were spent on fuel by the
electricity generators.
The following table summarises the overall electricity generation picture by type of fossil fuel in the USA
for 2007. This is based on the data in Reference [2] in which heat content of fuels is given on an HHV
basis, and the LHV efficiencies are derived therefrom by assuming average chemical properties of the
various fuel types.
Overall average efficiency, % (HHV)
Overall average efficiency, % (LHV)
Average fuel cost, c/kWh generated
1.2
Coal
Petroleum
Natural
Gas
32.8%
35.0%
1.32
31.4%
33.3%
11.7
39.5%
43.8%
6.14
Fossil
Fuels
Average
36.2%
39.0%
3.00
THE LAWS OF THERMODYNAMICS
Thermodynamics is a subtle science. The underlying concepts of Heat, Temperature and Work, are
physically tangible but not necessarily obvious. People feel "hot" and "cold". The distinction between
Heat, a quantity of energy, and Temperature, best described as the level of the heat energy, is subtle. To
this day, many perfectly intelligent people do not distinguish between Heat and Temperature. Likewise,
people know what Force and Distance are, so if you say that Work = Force x Distance the definition is
tangible; otherwise, try to explain in more fundamental terms what Work is ! (Let alone that it is a form of
energy, just like Heat).
Many brilliant scientists experimented with Thermodynamics before it was finally understood in a
quantitative sense and its fundamental laws established. Seventeenth century scientists, including Galileo
and The Florentine school in Italy, and Jean Rey in France, built thermometers to measure "heat" and
"coldness". They were measuring a quantity not yet defined, actually temperature, but no distinction
between "Heat" and "Temperature" had yet been made.
In the late 18th century, due to experiments in calorimetry by Joseph Black at Glasgow, the distinction
between heat and temperature began to be understood. The "Caloric Theory" stated the principal of
conservation of heat. Heat was described as consisting of "caloric", which is "a subtle, all-pervading
fluid". Caloric was thought to be stored in matter, and to have flowed from a hotter to a colder substance
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 M. A. Elmasri, 1990-2008
Chapter 1: Introduction
when they were brought into contact. Fire was a manifestation of caloric leaking out of the wood, and the
wood was exhausted and reduced to ashes when all its caloric had leaked out. This theory ran into
difficulty when trying to interpret the conversion of work into heat, since work was still not understood as
a form of energy, just like heat. Thus, when a cart's axle was heated by friction with the rotating cart
wheel, caloric theory tried to explain the phenomenon as due to leakage of caloric out of the materials,
due to the rubbing motion wearing off the materials, enabling the caloric to escape from them. The
obvious flaw of this explanation, that if you went on rubbing the same piece of matter, it kept on "leaking
caloric", as though it contained an infinite amount of the stuff, was obvious to Rumford and Joule,
amongst others. Despite the fundamental flaw of failing to acknowledge that work was a form of energy,
convertible to heat, caloric theory was accepted in the late 18th century and remained more or less current
until the mid 19th century.
Finally, in the mid 19th century, James Prescott Joule experimentally established the "Mechanical
Equivalent of Heat". In his landmark experiment, he put about an ounce of lead bird-shot into a long
glass tube, about 4 ft. (1.2 m) long and an inch (25 mm) in diameter, then plugged its ends with stoppers.
He then turned the tube upside down about a hundred times, causing the lead shot to fall along the length
of the tube and come to rest at its bottom with each turn. Thus, in a hundred turns, the lead would have
fallen about 400 ft (120 m). This warmed the lead shot by a few degrees, and it was then dumped into a
beaker containing a small, known amount of water, at a known temperature. The lead was then stirred it
into the water with a thermometer, which also measured the temperature rise of the water when brought
into equilibrium with the warm lead. This temperature rise was then used to calculate the amount of heat
generated within the lead by its falling a known distance. The experiment was repeated with different
amounts of lead, different tube lengths, and different numbers of turns of the tube. Thus, Joule found that
a BTU of heat was equivalent to 778 ft-lbs of mechanical work, effectively stating the First Law of
thermodynamics.
In simple language, the First Law of thermodynamics says that energy is conserved, cannot be created nor
destroyed, but can be converted from one form to another. While establishing the equivalence of Heat
and Work as forms of energy, the first law does not mention the restrictions on transforming energy from
one form to the other.
Although the principal of Conservation of Energy was embodied in 18th Century Caloric Theory, this
lacked the consideration of Work as a form of Energy. Thus the First Law of Thermodynamics,
expressing conservation of energy in its various forms, was not laid down until the mid-19th century by
Joule, Kelvin and Clausius. This was after steam locomotives were well developed. In 1845, the "Great
Western" steam locomotive had pulled a 100 ton train at an average speed of 59 mph between London
and Swindon.
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 M. A. Elmasri, 1990-2008
Chapter 1: Introduction
The Second Law of thermodynamics states the restrictions on converting energy from one form to
another. Simply put, the Second Law says that it is possible to totally convert Work into Heat, but only to
partially convert Heat into Work. More philosophically, we are always losing the battle, total chaos can
result from order but only partial order from chaos.
The portion of heat that can be converted into work depends on the temperature level of the heat. More
precisely, on the ratio of absolute temperatures between the heat source and the surroundings (heat sink).
All the heat could conceivably be converted to work if the source temperature was infinite or the sink
temperature absolute zero. Neither condition can exist in reality.
The Second Law was only stated completely by Planck and Poincare at the end of the 19th century and
the early years of the 20th. Many of its underlying concepts were part of Carnot's work of the early 19th
century and Clausius's work in the mid 19th century. However, Carnot's concept of a cycle, in which a
substance returned to its initial state after a series of processes, as well as his concept of a reversible
process, did not include Work as an energy interaction with his Caloric. Nonetheless, tradition has named
the maximum efficiency of a heat-to-work cycle the "Carnot Efficiency" in his honour.
Exergy is a thermodynamic property that measures the ability of a substance to produce work. It
comprises a chemical component, bound within the molecular structure of the substance, and a thermomechanical component, manifested in the sensible, non-chemically-reacting properties of pressure,
volume and temperature. The second law of thermodynamics implies that the exergy can at best be
preserved but tends to be destroyed in any real process.
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 M. A. Elmasri, 1990-2008
Chapter 1: Introduction
1.3
CONVERTING HEAT TO WORK – BASIC PRINCIPLES OF THE HEAT ENGINE
Heat cannot lift or move objects, unless it can be converted to Work. A Heat Engine is a device to effect
that conversion. Electricity is energy in the form of pure work, and is a convenient medium for the
transmittal of Work to the point where it can be used. A thermal power plant converts Heat to
Mechanical Work, i.e. Force x Distance, then converts that work to its electrical form for ease of
transmittal to its end-users.
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 M. A. Elmasri, 1990-2008
Chapter 1: Introduction
We can think of energy as having two basic forms: Heat and Work. Heat is energy that we normally
associate with temperature, which defines its level.
Work is energy that we normally associate with mechanical ability to push something, i.e. force
multiplied by distance.
Work, on the other hand, is organised, well-directed energy; uniform, smooth and highly convenient to
transmit and use. It can move things. Electricity is pure work.
One practical illustration of the difference between transporting heat and work is to consider a modern
combined cycle. Transferring 100 MW as thermal energy in the heat recovery boiler needs approximately
10 acres of heat transfer surface. 100 MW of work can be transmitted through a few square inches of
rotating shaft or electric conductor.
A heat engine, or thermal power system, converts heat, usually obtained by burning a fuel, to work. This
is done through the use of a "thermodynamically-coupled" working substance, i.e. a material whose
pressure x volume product increases when heated, such as steam or gases. Thus the ability of the working
substance to push on a piston or emerge at high velocity from a nozzle is increased by heating it. The
mechanical component, such as pistons or turbine blades, then convey the force of the working substance
to shafts transmitting the Work to its useful purpose. After pushing the mechanical component, the
exhausted working substance is then dumped or recycled.
It is impossible to construct a thermal power system using an uncoupled substance, such as an
incompressible liquid, since heating it will not increase its ability to "push".
1.4
1.4.1
THERMODYNAMICS OF POWER GENERATION
CONVERSION OF FUEL TO WORK:
Common fuels have a chemical exergy roughly equal to their Lower Heating Value (LHV).
Theoretically, their ability to do work is therefore approximately the same as their heat content. If
apparatus could be devised to convert their chemical potential to work we would have power systems that
are about 100% efficient. Unfortunately, direct conversion of chemical energy to work, such as via fuel
cells, has many practical difficulties and inefficiencies, so that working fuel cells are only about 50%
efficient. It may be noted that humans and animals convert the energy in carbohydrates into Work by a
fuel-cell-type reaction in the muscles. The efficiency of converting fuel energy to work in human
muscles is about 25%.
In thermal power systems, we burn the fuel to generate heat, then use the heat to increase the thermomechanical exergy of a working substance. The mechanical component of the exergy is then used to push
against an object, such as a piston or turbine blade, which delivers the work.
Because combustion of a fuel occurs at a finite temperature, it is irreversible and destroys part of its
exergy. As an extreme example, consider igniting a can of gasoline in an open space. Only a small
impetus is needed (a match) to start a spontaneous exothermic reaction that results in CO2 and H2O vapor
being formed from the gasoline and air. In a few minutes, all the heat released would have "disappeared"
into the surrounding air. The heat is still there and the surroundings are marginally warmer than they
were before combustion. From experience we know how irreversible this process is, since the
marginally-warmer air, CO2 and H2O vapor cannot spontaneously combine to re-create the gasoline and
marginally cool down the surroundings.
The above is an example of the ultimate irreversibility, destroying all the exergy of that gasoline without
accomplishing a desirable task. If that gasoline were burnt under pressure in the cylinder of a car, at least
we could have propelled the car for a few miles using part of the fuel's exergy captured as work by the
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 M. A. Elmasri, 1990-2008
Chapter 1: Introduction
pistons. The end result of the CO2, H2O vapor and warmer surroundings is the same, but in the process
we extracted useful work, which we used for a desirable task (such as driving to the mall !).
Most of the efficiency battle in a heat engine is lost at the first step of burning the fuel. Even if 100% of
its energy is transferred to the working substance, that is in the form of heat at a finite temperature, and
can only be partially converted to work. The exergy transferred to the working substance is therefore
considerably less than the exergy of the fuel.
If an amount of heat Q were transferred to a working substance at a constant temperature T, the associated
exergy transfer to the substance = Q (1 - Ta/T) , where Ta is the temperature of the surrounding ambient,
the ultimate heat sink for any cycle. This exergy transferred to the working substance is the maximum
amount of the energy Q that could possibly be extracted from that working substance by perfect,
reversible machinery. The remainder of the energy Q, i.e. Q (Ta/T) cannot possibly be converted to work,
no matter how perfect the processing of the working substance; and must leave the cycle in the form of
heat, not work. If our processing machinery is imperfect, as it surely is, the amount of the energy Q that
will have to leave the cycle as heat, not work, is even greater than Q (Ta/T), so the amount that can leave
as work is even less than Q (1 - Ta/T).
The above, simple relationships are for a working substance that receives heat at a constant temperature
T, such as water boiling into steam. If the working substance temperature increased as it received the heat
Q, a suitably averaged value of T should be used in the above expressions. If T increased linearly with Q
received (i.e. a working substance with constant heat capacity such as a perfect gas), the appropriate
average T can be shown to be the logarithmic mean of the temperatures at the beginning and end of the
heat addition process.
It is interesting to note that the fuel-cell conversion of carbohydrate energy to work by human muscles, at
about 25% efficiency, would have required a source temperature of at least 232 °F or 112 °C had it been
accomplished through a heat engine cycle. Fortunately, that reaction occurs at only 2 or 3 °F above body
temperature of 98.6 °F or 37 °C.
1-10
 M. A. Elmasri, 1990-2008
Chapter 1: Introduction
The value for Ta on earth varies with location and season, but the ISO definition for standard calculations
is set at the representative value of Ta = 518.67 °R (59 °F) = 288.15 K (15 °C).
Increasing the efficiency of thermal power systems can therefore be accomplished on two distinct fronts:
1. Maximise thermo-mechanical exergy delivered to the working substance by reducing the loss of exergy
(Q Ta/T) in combustion & heat transfer to the working substance. This means minimizing Ta/T. Since
there is nothing we can do to reduce Ta, we can only strive to increase T. This means transferring as
much heat as possible to the working substance at its highest possible temperature.
2. Maximise the conversion of the working substance's thermo-mechanical exergy to work output by
reducing the exergy losses in processing the working substance. This means efficient machinery and
processes to extract the exergy from the working substance.
Efficient processes in a thermal power system preserve exergy of the working substance by minimizing
losses of its exergy, i.e. pressure and temperature. Pressure represents the capability of doing work
directly (pv, pressure x volume = work). Temperature represents the capability of converting heat energy
to work. An example of a process that partially destroys the pv component is throttling a steam flow. An
example of one that partially destroys the thermal component is transferring heat from a hot to a cooler
substance (e.g. exhaust gas to steam). Unfortunately, such processes are necessary in practice.
Efficiency-degrading processes can always be logically identified because they are irreversible. A stream
of high pressure steam expanded across a reducing valve will spontaneously flow into the lower
downstream pressure. Heat will spontaneously flow from hot to cold fluid in a heat exchanger. Neither
can spontaneously flow back. We can restore the steam to the higher pressure (by a compressor) or the
heat to a higher temperature (by a heat pump), but in so doing, we would have to expend work.
Measures to reduce irreversibility will improve efficiency, but will also increase capital cost. For
instance, larger diameter pipes and valves will lower the pressure losses in a steam plant, reducing exergy
losses and improving efficiency. Those pipes and valves, however, are more expensive than ones of
smaller diameter. Likewise, a multi-pressure heat recovery boiler has a smaller overall temperature
difference between flue gases and steam compared to a single pressure boiler. This reduces exergy lost in
1-11
 M. A. Elmasri, 1990-2008
Chapter 1: Introduction
heat transfer from gases to steam but requires a larger surface area and is more complex and expensive.
Understanding tradeoffs of that sort is the key to good plant design.
1.4.2
ENERGY & EXERGY BALANCE OF A CONVENTIONAL STEAM PLANT:
Fig. 6 shows typical energy and exergy balances for a conventional, subcritical, single-reheat steam
power plant burning clean fuel with minimal parasitic losses. The conditions are 2700 psia/1040 °F (186
bar/560 °C) with reheat at 540 psia/1040 °F (37 bar/560 °C) and condenser at 0.65 psia (.045 bar). The
cycle has 8 heaters with the top one fed from the HPT above the reheat point (so-called HARP cycle).
The stack temperature of this plant is 266 °F (130 °C). The plant has a gross efficiency at generator
terminals of about 44.3 % (LHV) and a net efficiency, after auxiliaries, of about 42.5% (LHV). This is
about as efficient a steam plant as can be built at ISO conditions with some measure of economy.
Supercritical plants with double reheats and ten heaters can attain up to 46% net efficiency, if maximum
efficiency is sought and the additional cost and complexity were acceptable.
1-12
 M. A. Elmasri, 1990-2008
Chapter 1: Introduction
Misc 1
Work 42.5
Misc 1
Cond 2.5
Work 42.5
Stack 5
Turbine 6
Boiler 2
Stack 1
Cond 49.5
Boiler 47
Energy Balance
as % of Fuel LHV
Exergy Balance
as % of Fuel LHV
Fig. 6. Energy & Exergy Balances for a Conventional 380 MW Steam Power Plant, clean fuel, 8 FW
heaters, 1 reheat
F
Q
o
o
C
Heat Addition in a Steam Cycle
1200
600
1000
500
T
400
300
600
200
400
100
200
0
Teq = approx. 770 oF (410 oC)
800
0
S
Ωlost = Ta / Teq = 42%
1-13
 M. A. Elmasri, 1990-2008
Chapter 1: Introduction
The energy balance of Fig. 6 shows almost 50% of the fuel's heat being lost at the condenser, 5 % at the
stack, and only about 2 % in the boiler, including blowdown and heat losses from the boiler itself as well
as pipes between boiler and turbine.
The energy balance shows where the energy goes but not why. For example, half the heat is lost at the
condenser. Does that mean that designing a better condenser would significantly reduce this loss ?
Obviously not ! That enormous condenser heat loss was created by other cycle processes. As a matter of
fact, most of the cycle's inefficiency was created before the fuel's energy reached the steam in the first
place. This is shown by the exergy balance. Almost half the fuel's capability to produce power was
destroyed by combustion and heat transfer because the "grade of heat" transferred to the steam is far
below the fuel's potential. By contrast, the condenser loses very little exergy. The heat rejected from it is
of such a low grade, that an infinitely large condenser allowing the steam to condense exactly at ambient
temperature would only salvage 2.5 % more power.
For our typical steam power plant, the maximum working substance (live steam) temperature at the start
of work extraction is 1040 °F (560 °C). The boiler generates this steam from feedwater leaving the FW
heaters at 544 °F (285 °C). The equivalent mean temperature of the water & steam receiving heat in the
economiser, evaporator, superheater and reheater is about 770 °F (410 °C). Thus, the "heat degradation
loss" in heat transfer from gases to steam is about 42%
(59+460)/(770+460) = .42
(English)
(15+273)/(410+273) = .42
(Metric)
In addition, we have thermal degradation losses in the air preheater as well as exergy losses associated
with pressure drops in the boiler and piping to and from the steam turbine, as well as the various "minor
heat losses", so we end up losing about 47% of our exergy in the boiler.
The efficiency of steam plants can be improved by raising the feedwater temperature, raising the pressure
(to raise saturation temperature in the evaporator) or raising live steam temperature. All those parameters
would tend to raise the mean temperature of heat reception by the water & steam at the high pressure.
Likewise raising the reheat temperature or employing double reheats in conjunction with high
(supercritical) pressures. With considerations of practical physical construction limiting pressures to
about 4500 psia (310 bar) and steam temperatures to about 1075 °F (580 °C), we find that those
constraints condemn the pure steam power plant to losing over 40% of the fuel's exergy to thermal
degradation in transferring heat into its working fluid.
The losses in turbines, generator, pumps, piping, and condenser will inevitably cost an additional 11 or
12% of the fuel exergy for a maximum net efficiency of about 46%.
1.4.3
ENERGY & EXERGY BALANCE OF A GAS TURBINE & COMBINED CYCLE:
For gas turbines, the maximum working substance temperature at the start of work extraction is about
2400 °F (1315 °C), for late 1990’s technology. The mean temperature of heat addition to the
air/combustion gases working substance, between compressor discharge, about 740 °F (395 °C) and
turbine rotor entry is about 1450 °F (790 °C). (Logarithmic mean, not arithmetic mean, should be used
for averaging the absolute temperatures). About 73% of the fuel exergy is transferred to the working
substance.
1 - (59+460)/(1450+460) = .73
(English)
1 - (15+273)/(790+273) = .73
(Metric)
That percentage can be increased by raising the mean temperature of heat addition, either by raising
pressure ratio (raising the air inlet temperature to the combustor), raising firing temperatures, or
employing reheat (staged combustion).
1-14
 M. A. Elmasri, 1990-2008
Chapter 1: Introduction
Fig. 7 shows typical energy and exergy balances for a modern gas turbine. Transferring the fuel's exergy
to the high-temperature working substance (combustion gases) destroys only 27 % of the exergy.
GT Work 37
GT Work 37
Misc 1
Compressor 2
Turbine 5
Exhaust 29
Exhaust 62
Combustor 27
Energy Balance
as % of Fuel LHV
Exergy Balance
as % of Fuel LHV
Fig. 7. Energy & Exergy Balances for a Modern Heavy Duty Gas Turbine
PR = 16, TRIT = 2400 °F = 1315 °C
Compared to conventional steam systems, gas turbine based systems possess an inherent advantage on the
first front (reducing loss in exergy transfer to the working substance). They are at some disadvantage on
the second front (efficient processing of the working substance) since more exergy is destroyed after the
initial transfer because of (a) necessity of cooling the hot working parts of the turbine, and (b) the
unwieldly process of recovering exergy from the exhaust gases, due to their large massflow at modest
temperature.
Nevertheless, the inherent advantage on the first front is so large that a typical large combined cycle of
the late 1990’s provides about 56% net ISO efficiency compared to about 43% for conventional steam
cycles. Fig. 8 illustrates energy and exergy balances for a combined cycle of 55.5 % net ISO efficiency.
Fig. 9 shows efficiencies of some GE gas turbines, based on data current as of writing (Fall, 2005), with
4” (10 mb) inlet and 2” (5 mb) exhaust pressure losses. Electricity efficiency is the ratio of power output
at generator terminals to fuel LHV energy input. Exergy efficiency includes the exhaust gas exergy in the
numerator as well. Thus, if a hypothetical, perfect bottoming cycle were possible, the resulting combined
cycle would be at that exergy efficiency, which thus represents the theoretical upper bound on the
efficiency of a combined cycle based on each gas turbine model. Please note that for the LMS100, the
heat extracted from the compressor intercooler is assumed lost in this calculation, as it is useful for
cogeneration but of very little value for a pure power generation combined cycle.
1-15
 M. A. Elmasri, 1990-2008
Chapter 1: Introduction
1400
F
Q
o
o
C
Heat Addition in a Gas Turbine
2800
2400
1200
2000
T
1000
800
600
1600
Teq = approx. 1450 oF (790 oC)
1200
400
800
200
400
0
0
S
Ωlost = Ta / Teq = 27%
GT Work 37
GT Work 37
ST Work 20
Misc 1
ST Work 20
Compressor 2
Turbine 5
Stack 1.5
H. T. 3.3
Cond 1.6
ST 2.6
Stack 9
Cond 23
Combustor 27
Energy Balance
as % of Fuel LHV
Exergy Balance
as % of Fuel LHV
Fig. 8. Energy & Exergy Balances for a typical 380 MW Combined Cycle, 1 GT + 1 ST, Net ISO
Efficiency 55.5%
1-16
 M. A. Elmasri, 1990-2008
Chapter 1: Introduction
Fig. 9. Efficiencies of Some Gas Turbines (ISO conditions, CH4 fuel, with typical inlet & exhaust losses.
GT data as of late 2005. Exergy reference 15 ºC.)
1.4.4
THERMODYNAMICS OF COGENERATION
1.4.5
EXERGY EFFICIENCY OF PROCESS HEAT FROM FUEL
Many applications require heat at relatively low temperature. Examples are heating a building or drying
an industrial product. When fuel is burnt solely to satisfy those needs, an irreversible loss of workproducing capability (exergy) occurs because the fuel could have supplied heat at a much higher
temperature than the process requires. The high-temperature heat could have produced work and the
residual, low-temperature exhaust heat from the engine could have been used for the process.
If methane (CH4) is burnt with 15% excess air, starting from ambient conditions, the products of
combustion would be produced at the adiabatic flame temperature, about 3300 °F (1600 °C). Those
products of combustion would posses virtually 100% of the fuel's energy but only 80% of its exergy.
Cooling the products of combustion in a boiler to make steam, would recover 93% of their energy
(assuming a stack temperature of 250 °F (121 °C) and a 1% heat loss). If that boiler produced saturated
steam at 1 atmosphere from ambient feedwater, the exergy recovered would only be 20% of the fuel's
exergy. This is because the steam we are producing, which may be used to heat a building, is a lowexergy product.
1-17
 M. A. Elmasri, 1990-2008
Chapter 1: Introduction
If the above calculation were done for a boiler producing 100 psia (6.9 bar) saturated steam, the exergy
efficiency would be about 28%. If the boiler were producing 500 psia (34 bar) saturated steam, the
exergy efficiency would be 35%.
Thus burning fuel to produce heating steam is only 20% exergy-efficient for low pressure steam (such as
used to heat buildings) and about 30% exergy-efficient for medium pressure steam (such as used for many
industrial processes). Burning fuel to produce modest-grade heat is a waste of its exergy. That fuel is
capable of doing better things, such as generate power, a more useful and valuable commodity. Since
generating power always involves rejecting heat, it stands to reason that cogenerating power and heat is
the most effective way of using fuel, where the high-temperature, high-exergy part of its energy makes
power and the remaining, low-temperature, low-exergy part of its energy is used for heat.
1.4.6
EFFICIENCY DEFINITIONS
As an example, consider an industrial process requiring 100 units of heat in the form of saturated steam at
a temperature of 212 °F (100 °C) generated from water at ambient temperature. Fig. 10 shows three
options of supplying that steam.
If the steam was generated by burning fuel in a 93 % efficient (LHV) boiler, 108 units of fuel would be
used. While the energy efficiency is 93 %, the exergy efficiency is only 20%. Fig. 11 shows the energy
and exergy balances for the boiler.
If instead of burning the fuel in a boiler, we burnt it in a typical modern gas turbine sized to allow the 100
units of steam to be generated in an exhaust heat recovery boiler, the gas turbine would consume about
180 units of fuel energy and produce 60 of electric output and 120 of exhaust heat, of which 100 would be
recovered.
Efficiency Definitions
For a pure power plant
η = Electricity Output
Fuel Input
For a Cogeneration Plant, there are two products: Power & Heat:
ηtotal = Electricity Output + Useful Heat Output
Fuel Input
ηPURPA = Electricity Output + ½ Useful Heat Output
Fuel Input
ηchargeable =
Electricity Output
Fuel Chargeable to Power
Chargeable fuel = Actual fuel – Fuel that would otherwise be used just to make
the same steam in a boiler, without cogeneration
1-18
 M. A. Elmasri, 1990-2008
Chapter 1: Introduction
Process Steam
100
Fuel
108
Electricity
60
The GT size is
such that it can
provide 100 units
of recoverable
exhaust energy
ηT = 93%
ηΩ = 21%
Package Boiler
Fuel
180
GT
Process Steam
100
Heat Recovery Boiler
Electricity
130
ηT = 89%
ηΩ = 45%
ηch = 83%
Process Steam
100
ST
The GT size is
such that useful
heat output is just
over 15% of total
useful energy
output
Electricity
290
Fuel
850
GT
HP
IP
Condenser
Heat Recovery Boiler
Fig. 10. Three Methods of Obtaining 100 Energy Units of Cogeneration Steam
1-19
ηT = 61%
ηΩ = 52%
ηch = 57%
 M. A. Elmasri, 1990-2008
Chapter 1: Introduction
Loss 7%
Loss 79%
Steam 93%
Steam 21%
Energy %
Exergy %
Fig. 11. Energy & Exergy Utilisation, Simple Boiler making 1 atm. stm
Loss 11%
Elect 33%
Loss 55%
Steam 56%
Elect 33%
Steam 12%
Energy %
Exergy %
Energy efficiency = 89 %
Exergy efficiency = 45 %
Chargable efficiency = 83%
Fig. 12. Energy & Exergy Utilisation, GT/HRB Cogen Plant making 1 atm. stm
1-20
 M. A. Elmasri, 1990-2008
Chapter 1: Introduction
Crude energy utilisation efficiency of cogeneration plants is usually measured by the "Total Efficiency"
or "CHP (Combined Heat and Power) Efficiency"
( Electricity Output + Total Heat Output ) / Fuel Input
The total energy efficiency of the above is 160/180 = 89%, i.e. below that of a simple boiler. That this is
a crude measure of the efficiency is evidenced by the fact that it makes no distinction between the value
of electricity (Work) and the value of Heat. The exergy efficiency of the above plant is 60+21/180 = 45
% or more than double that of the simple boiler. Fig. 12 shows energy and exergy balances.
Recognizing that electricity is more valuable than heat, the U.S. congress passed the "Public Utility
Regulatory Policy Act" or PURPA legislation in 1979. This uses a measure of cogeneration plant
efficiency that assigns to heat output half the weight of electric output, i.e.
PURPA Efficiency = ( Electricity Output + 1/2 Heat Output ) / Fuel Input
PURPA efficiency of the above example is (60+ 50)/180 = 61 %.
Yet another measure of cogeneration plant efficiency is the ratio of its net electricity output to the excess
fuel it consumes, over and above the fuel that would have been used to generate steam in a boiler, without
cogeneration:
Chargeable Power Generation Efficiency
= ( Electricity Output + Boiler Electricity Saved ) / ( Fuel Consumed - Boiler Fuel Saved )
Sometimes the boiler electricity saved is ignored to be on the conservative side in quoting that efficiency.
This is the case in the present example where fuel consumed by the gas turbine, over and above that by
the simple boiler (180-108=72 units) is charged to the electric output of 60 units. Thus the Chargeable
Power Generation Efficiency is 60/72 = 83 %; far above what can be achieved by any pure-power, noncogeneration plant.
The electricity-to-heat ratio of the above example is 0.6, suitable for in-house consumption of all
electricity and heat produced by many industrial plants.
The third option illustrated in Fig. 10 is an "Electrically-Oversized" system, a 3-pressure condensing
combined cycle with the required process steam bled from the steam turbine and/or boiler. The use of a
condensing steam turbine allows a broad range of reasonable plant sizes, since any amount of steam
generated in the heat recovery boiler in excess of the process requirement can be used to produce more
power. A typical design of such a system, sized for 100 units of steam "by-product" would produce about
420 units of electricity and consume 850 units of fuel. Its energy efficiency is 520/850 = 61 %, well
below that of the second option. Its exergy efficiency 420+21/850 = 52 %, above that of the second
option. Its electricity-to-heat ratio is 4.2, too high for in-house consumption by most industries and likely
to need power export to the grid.
Fig. 13 shows energy and exergy balances for the third option. The fuel consumed by the gas turbine,
over and above that by the simple boiler, is 850-108=742 units. The Chargeable Power Generation
Efficiency is therefore 420/742 = 57 %, good but well below that for the second option.
1-21
 M. A. Elmasri, 1990-2008
Chapter 1: Introduction
Loss 39%
Loss 48%
Elect 49%
Elect 49%
Steam 12%
Steam 3%
Energy %
Exergy %
Fig. 13. Energy & Exergy Utilisation, 3-P Cond CC with cogenerated 1 atm. stm
1.4.7
ECONOMIC CONSIDERATIONS
Thermodynamics and economics are intimately related. Because work is more useful than heat, it is more
valuable (and more expensive). Electricity costs typically 3 to 6 times as much as fuel on an energy basis.
To make electricity, not only must we use more fuel than the electricity produced, we also have to expend
capital on plant equipment to carry out the conversion.
Likewise, steam is worth more per unit of its heat content than fuel, since to generate steam in a boiler,
we must burn more fuel on an energy basis (about 7% more at 93% boiler efficiency) and we must also
buy the boiler and spend money operating and maintaining it.
Thus, if fuel thermal energy on an LHV basis is worth 1 unit of money per unit of energy, low pressure
process steam is typically worth 1.5 units of money per unit of energy and electricity is typically worth 5
units of money per unit of energy.
Applying those numbers to the above three examples, we get
Simple boiler (Option 1):
Value of Input Fuel = 108 x 1 = 108
Value of Output Steam = 100 x 1.5 = 150
Value added = 150-108 = 42
Value Out / Value In = 150/108 = 1.4
Cogeneration Plant (Option 2):
Value of Input Fuel = 180 x 1 = 180
Value of Output Steam = 100 x 1.5 = 150
Value of Output Electricity = 60 x 5 = 300
Value added = 300+150-180 = 270
Electricity Value / Chargeable Fuel Value = 300/72 = 4.2
Value Out / Value In = 450/180 = 2.5
Cogeneration Plant (Option 3):
1-22
 M. A. Elmasri, 1990-2008
Chapter 1: Introduction
Value of Input Fuel = 850 x 1 = 850
Value of Output Steam = 100 x 1.5 = 150
Value of Output Electricity = 420 x 5 = 2100
Value added = 2100+150-850 = 1400
Electricity Value / Chargeable Fuel Value = 2100/742 = 2.8
Value Out / Value In = 2250/850 = 2.65
p
2400
2250
2200
2000
Economic Value
1800
1600
1400
1200
1000
850
800
600
450
400
200
108
150
180
Steam
Fuel
0
Fuel
Stm/Elect
Fuel
Stm/Elect
Fig. 14. Typical Relative Economic Values for a Simple Boiler, GT/Boiler, and Cond/Extr CC with the
same steam output
Fig. 14 shows the comparative economic values of the inflowing and outflowing commodities for each of
the three plant options. Naturally, the cogeneration plants will cost much more capital than the simple
boiler and will be significantly more costly to operate and maintain. Option 3 would be more complex
than Option 2 but will benefit from a relative economy of scale.
Cogeneration is less energy-efficient than a boiler. That is true locally, at the cogeneration site, but not
globally, considering the replacement of the cogenerated electricity by a conventional power plant. With
Option 2, had the 60 units of electricity from the cogeneration plant come from a 55% efficient poweronly combined cycle, that plant would have burnt an additional 109 units of fuel. Thus the amount of fuel
burnt globally without cogeneration would have been 108+109 = 217 versus the 180 burnt with
cogeneration. In the global sense the cogeneration plant has saved 37 units of fuel. For Option 3, the 420
units of electricity from a power-only combined cycle would have consumed 420/.55 = 764 units of fuel
at the central station, so global fuel consumption would have been 108+764=872 vs 850 with
cogeneration, for a fuel saving of 22 units with cogeneration. Thus, on a global basis, Option 2 is better
than Option 3.
In summary the example given above illustrates the following general rules of the Thermodynamics of
cogeneration:
1-23
 M. A. Elmasri, 1990-2008
Chapter 1: Introduction
1. A cogeneration plant is usually less energy-efficient but more exergy-efficient than a simple boiler.
2. Cogeneration usually reduces local energy efficiency but increases global energy efficiency.
3. Cogeneration always increases exergy efficiency, locally as well as globally.
4. The more energy efficient options are usually less exergy efficient and vise versa.
5. The laws of economics and thermodynamics are inseparable. They dictate that the more-refined
product (exergy, work, electricity) has a much higher economic value than the raw material (fuel) and the
less-refined product (steam in the example). Naturally, producing the high-value product requires more
sophisticated equipment which costs more capital.
Although cogeneration will always save fuel from the global viewpoint, much of our heat consumption,
such as heating houses, is dispersed. Efficient electricity generation requires complex plant that benefits
from economy of scale, favouring concentrated generation of electricity. Distributing electricity is easy
but distributing heat is not. Work, the organised energy, is very easy to transmit, a few square inches of
transmission line conductors can carry hundreds of MW over hundreds of miles with a loss of only a few
percent. Heat, the disorganised energy, cannot be transmitted in a compact, economical, low-loss fashion
except for relatively short distances of up to a few miles. Thus, cogeneration is most suitable for
situations where a heat load is concentrated close to a large demand for power or to a site suitable for a
large power plant. Examples are district heating for cities or towns with high population density and
industrial sites which have high heat demands.
In 2007 in the USA, about 7 Quads were used to heat houses, and 4 Quads were used to heat commercial
buildings; i.e. about 11 Quads of heat was used in the USA at very modest temperatures, about 100 °F (38
°C) at its end-use point. The heat rejected from the condensers and stacks of all thermal power plants
(fossil and nuclear) was about 25 Quads, more than twice the heat needed by all houses and commercial
buildings in the country. Unfortunately, heat cannot be recovered and transported efficiently and costeffectively from all the power plants in the country to all the buildings in the country. Had it been, we
would have saved about $130 billion worth of heating fuel bills in 2007.
In the USA, under 5% of the electricity generated and the heat consumed is from cogeneration. Although
cogeneration cannot supply all our thermal needs, it can certainly be expanded from these modest
numbers to yield substantial savings.
References:
[1] Key World Energy Statistics, 2007. International Energy Agency (IEA).
[2] Annual Energy Review, 2007. Energy Information Administration (EIA), US Department of Energy.
1-24
 M. A. Elmasri, 1990-2007
Chapter 2: GT Cycle Thermodynamics
GAS TURBINE CYCLE FUNDAMENTAL THERMODYNAMICS
Revised October, 2007
© Maher Elmasri 1990-2007
2.1
INTRODUCTION
The gas turbine cycle is introduced and its fundamental thermodynamics described. The ideal and real
cycle models are presented.
Design-point analysis, for generic engines with typical component assumptions, is used to find
performance of gas turbine cycles in various configurations and arrangements. This provides insight into
the effect of key cycle parameters on gas turbine thermodynamic performance, and allows comparison of
the characteristics of various cycle types. In the discussions and charts of this chapter, each point
corresponding to a set of cycle thermodynamic parameters represents a machine designed to operate with
the specified thermodynamic parameters at its nominal operating condition. This is in contrast to the rest
of this seminar, where the discussion focuses on specific machines operating under variable conditions.
To keep the discussion simple, the results below are for pure power generation. Any of the cycles
discussed can be adapted for cogeneration in a variety of ways. Since the rest of this seminar considers
combined cycles in greater depth, their treatment in this chapter is incomplete, and only serves to relate
the effects of gas turbine cycle design parameters to combined cycle performance.
2.2
2.2.1
THE BRAYTON CYCLE
PROCESS DESCRIPTION
Figure 1 illustrates the basic components of a Brayton Cycle gas turbine, with a tabulation of pressure,
temperature, mass and volume flow rates through a typical heavy duty gas turbine of the 1990’s. It also
shows a T-s sketch of the cycle. The h-s diagram of the cycle is essentially similar.
The compressor draws in atmospheric air through an inlet filter system with a small relative pressure loss
(about 1%). The air is compressed (1-2) and delivered to the combustor where fuel is directly burnt. A
relatively small pressure drop and large temperature rise (2-3) occur through the combustor. The
resulting high temperature product gases have a much larger volume than the air entering the combustor,
therefore their p v product and ability to do work is much greater. Those expand through the turbine (34), producing more power than consumed by the compressor. The turbine drives the compressor and the
excess work overcomes mechanical losses and drives the load. The exhaust gases leaving the turbine are
at a pressure slightly higher than atmospheric but are still at an elevated temperature. Because the cycle
performs better at high turbine inlet temperatures, that cannot be endured by uncooled turbine
components, some air is bled from the compressor and used to cool the turbine.
In studying Fig. 1, it is worthwhile to note that the volumetric flow rate entering the combustor is 13%,
whereas that leaving it is 33% of compressor inlet volumetric flow rate. It is this increase in volumetric
flow rate, by a factor of about 2½, at high pressure, that enables the turbine to produce greater power than
consumed by the compressor. With isentropic turbomachines, zero pressure losses, and constant mass
flows and fluid properties throughout, the turbine’s work in this case would be 2½ times the
compressor’s. In a typical gas turbine, with all real effects, the turbine power output is roughly twice the
compressor's power consumption. The net power is thus about half that produced by the turbine.
2-1
 M. A. Elmasri, 1990-2007
Chapter 2: GT Cycle Thermodynamics
Atmospheric
Air
Fuel
1
2
4
3
C
T
Exhaust
Gas
G
Temperature
6
Cooling Air
3
5
Typical Heavy Duty GT
θ
Point
Pressure (atm)
Mass Flow
Volume Flow
Temperature θ
Temperature K
Temperature ºC
Temperature ºR
Temperature ºF
4
3
4
2
1
1
100
100
1
288
15
519
59
2
15
86
13
2.25
648
360
1168
710
3
14.5
88
33
5.5
1584
1296
2855
2396
4
1.03
102
300
3
864
576
1557
1098
2
1
1
s
Figure 1. Brayton Cycle configuration, T-s diagram, and typical values of pressure, temperature and mass flow rate
at its four cardinal points, all relative to the values at the engine inlet (State 1)
2.2.2
IDEAL, AIR-STANDARD, BRAYTON CYCLE ANALYSIS
For that simple model, we assume:
* The working substance is air throughout.
* Air is a perfect gas with constant properties.
* Fuel mass flow is negligible and combustion is equivalent to heat addition.
* Compressor and turbine are isentropic (100% efficient).
* Inlet, combustor and exit pressure losses are zero.
* Turbine is uncooled.
Fig. 2 shows such a cycle on the T-s diagram, where the scale on the y-axis is the dimensionless absolute
temperature, θ=T/T1. The curved lines 1-4 and 2-3 are constant-pressure lines. For the perfect gas model,
temperature increases exponentially with entropy along a constant-pressure curve, since the fundamental
relationship between thermodynamic properties,
T ds = dh – v dp ,
gives, for constant pressure and a perfect gas,
ds = Cp dT/T ,
and thus, by separation of variables and integration,
T = T0 e
( s − s0 )
Cp
2-2
.
 M. A. Elmasri, 1990-2007
Chapter 2: GT Cycle Thermodynamics
Because the constant-pressure curves steepen as entropy increases, the vertical distance between them is
larger at higher entropy. This vertical distance is proportional to the work of an isentropic (adiabatic,
reversible) compression or expansion between the two pressures. Thus, the expansion 3-4 produces
greater work than the compression 1-2 requires, and that difference is the cycle’s net work output.
τ=
T3 T2  P2 
=
= 
T4 T1  P1 
θ=
γ −1
γ
3
T
3-4: Isentropic
Expansion
2-3: Constant Pressure
Heat Addition
T3
T1
4
2
1-2 : Isentropic
Compression
1
4-1: Constant Pressure
Heat Rejection
s
Compressor work per unit mass:
WC = Cp (T2-T1) = CpT1(τ-1)
Turbine work per unit mass:
WT = Cp (T3-T4) = CpT1 θ (τ-1)/ τ
Net work per unit mass:
WNET = WT – WC = CpT1 (θ-τ) (τ-1) / τ
Heat added per unit mass:
QADD = Cp (T3-T2) = CpT1 (θ-τ)
Cycle efficiency:
η = WNET / QADD = (τ-1)/ τ
Figure 2. Ideal, air-standard model of a Brayton Cycle
The cycle has two independent parameters that the designer can choose, the pressure ratio and the
maximum temperature. The effect of these parameters on cycle performance can be easily calculated
algebraically in the air-standard cycle as shown below.
Let π = cycle pressure ratio = P2 / P1 = P3 / P4
............ (1)
For the isentropic compression 1-2 and expansion 3-4 the absolute temperature ratio τ is
x
τ = T2 / T1 = T3 / T4 = π ...................………. (2)
where
x = (γ−1)/γ = R/Cp (= 0.2857 for air)
γ = ratio of specific heat at constant pressure to that at constant volume (= 1.4 for air)
R = gas constant (= 53.4 ft-lbf/lbm-R = 0.0686 BTU/lb-°R = 287 J/kg-K for air)
Cp = specific heat at constant pressure (= 0.24 BTU/lb-°R = 1004 J/kg-K for air)
Thus for a unit mass of air the compressor work is:
WC = h2 – h1 = Cp (T2 – T1) = Cp T1 (τ - 1) ...…………. (3)
and the turbine work is
WT = h3 – h4 = Cp (T3 – T4) = Cp T3 (1 - 1/τ) .……….... (4)
Letting θ = T3/T1 = ratio of turbine inlet temperature to ambient temperature, the net work is
WNET = WT – WC = Cp T1 { θ (1 - 1/τ) - (τ-1) } ……..... (5)
2-3
 M. A. Elmasri, 1990-2007
Chapter 2: GT Cycle Thermodynamics
…. = Cp T1 (τ-1) (θ-τ) / τ ......….……………….…..... (5a)
Thus, we see that net work per unit of airflow, also called specific work or specific power, depends on
pressure ratio as well as dimensionless maximum temperature, θ. Since equipment size increases in
proportion to airflow, specific work bears a rough proportionality to the ratio between power output and
cost. Thus, it is desirable to maximise specific work in order to reduce equipment cost per unit of power
output, i.e. reduce $/kW of capital cost. Naturally, the design features that maximise specific work may
increase the cost of the equipment in other ways, so the tradeoff in practice is more complex.
The heat added in the idealised combustor is
QADD = h3 – h2 = Cp (T3 – T2) = Cp T1 (θ - τ) ………..... (6)
and the thermal efficiency of the cycle is
x
η = Wnet / Qadd = 1 - 1/τ = 1 - 1/π
..........………..... (7)
which depends only on the pressure ratio and has no dependence on turbine inlet temperature.
Fig. 3 shows the ideal cycle performance on efficiency vs. specific power coordinates, easily plotted from
equations 5 and 7. The horizontal lines represent pressure ratios from 6 to 42 as shown on the legend.
The curves are for different values of θ, T3/T1, the dimensionless value of T3. To render θ more
meaningful, the corresponding value of T3 in ºC at ISO conditions (T1=15 ºC = 288.15 K) is also shown
on the legend. For those accustomed to “Imperial Units”, the table below gives the corresponding
temperatures in ºF and ºR:
Table 1. Temperatures in different units and the corresponding dimensionless temperature θ referenced to ISO
conditions
°C
950
1000
1050
1100
1150
1200
1250
1300
1350
1400
1450
1500
1550
K
θ (relative to
288.15 K)
1223
1273
1323
1373
1423
1473
1523
1573
1623
1673
1723
1773
1823
4.24
4.42
4.59
4.77
4.94
5.11
5.29
5.46
5.63
5.81
5.98
6.15
6.33
°R
2202
2292
2382
2472
2562
2652
2742
2832
2922
3012
3102
3192
3282
°F
1742
1832
1922
2012
2102
2192
2282
2372
2462
2552
2642
2732
2822
Per equation 7, theoretical efficiency depends only on pressure ratio, and increases from about 40.1% for
a pressure ratio of 6 to 65.6% for a pressure ratio of 42. As equation 5 shows, the specific power is a
function of turbine inlet temperature, θ as well as pressure ratio. For each value of θ, the specific power
peaks at a certain pressure ratio. By setting the partial derivative of WNET with respect to τ to zero from
equation 5 one finds that the maximum specific power occurs when τ = √θ, i.e. at the pressure ratio
corresponding to the compressor, combustor, and turbine having equal temperature ratios, with exhaust
temperature (T4) equaling compressor discharge temperature (T2).
2-4
 M. A. Elmasri, 1990-2007
Chapter 2: GT Cycle Thermodynamics
Ideal Air-standard Brayton Cycle
70%
65%
PR=6
PR=12
Cycle Efficiency
60%
PR=18
PR=24
55%
PR=30
PR=42
θ=4.592 (1050 °C)
50%
θ=5.112 (1200 °C)
θ=5.633 (1350 °C)
45%
θ=6.154 (1500 °C)
Max specific power
40%
35%
250
300
350
400
450
500
550
600
650
700
Specific Power kJ/kg (kW per kg/s)
Figure 3. Performance of theoretical air-standard Brayton Cycle
2.2.3
IDEAL, AIR-STANDARD BRAYTON CYCLE ANALYSIS - SECOND LAW APPROACH
The above is the standard, textbook analysis to derive the performance of the Ideal Brayton Cycle. It is
instructive to re-derive it using an unconventional second law approach. To the assumptions given above
we add the idealisation that the exergy of the fuel equals its heating value. This is approximately true of
all practical fuels. Methane, for instance has a theoretical exergy just 0.7% greater than its LHV.
2.2.3.1 The Equivalent Temperature of the Variable-Temperature Heat Addition and Rejection
Consider a stream of fluid of massflow rate m and constant specific heat Cp, such as a perfect gas,
cooling down from an initial temperature Ti to a final temperature Tf. If the ambient temperature is Ta, an
increment of heat dQ extracted from it at temperature T can provide work
dW = dQ ( 1 – Ta / T ) ....................... (8)
in a perfect Carnot Cycle. This heat extraction reduces its temperature by an increment
dT = -dQ / ( m Cp ). ......................... (9)
Thus
dW = -m Cp ( 1 - Ta / T ) dT
......... (10)
Integrating from Ti to Tf gives the work that can be extracted as
W = -m Cp ( Tf – Ti ) + m Cp Ta ln ( Tf / Ti )
..…. (11)
which is the change in exergy between the initial and final states. The heat transferred from the fluid
between Ti and Tf is
Q = -m Cp ( Tf – Ti ) ......................... (12)
2-5
 M. A. Elmasri, 1990-2007
Chapter 2: GT Cycle Thermodynamics
If that fluid were replaced by a heat source at constant temperature Ts, giving the same amount of heat Q,
the work that could be generated in a perfect Carnot Cycle would be
Q (1 - Ta / Ts) = -m Cp (Tf - Ti) (1 - Ta / Ts)
......…... (13)
Combining 11 and 13 gives
Ts = ( Tf - Ti) / ln (Tf / Ti)
.......…... (14)
i.e. that the equivalent constant source temperature is the logarithmic mean of the initial and final
temperatures.
2.2.3.2 Representing the Brayton Cycle by its equivalent Carnot Cycle
Returning to the cycle of Fig. 2, we now see that the variable-temperature heat addition process 2-3 is
equivalent to one at a constant temperature of
Tadd = (T3 – T2) / ln (T3/T2)
........……........ (15)
and the heat rejection process 4-1 is equivalent to one at a constant temperature of
Trej = (T4 – T1) / ln (T4/T1)
.....……........ (16)
and that the ideal Brayton Cycle is equivalent to a Carnot Cycle between temperatures Tadd and Trej whose
efficiency is
η = 1 - Trej / Tadd ...........……............... (17)
This equivalent Carnot Cycle is shown on Fig. 4.
Noting from eq. 2 that τ = T2 / T1 = T3 / T4 , so that T4 / T1 = T3 / T2 , we see that equation 17 reduces to
η = 1 - 1/τ
which is the same result as equation 7.
2-6
 M. A. Elmasri, 1990-2007
Chapter 2: GT Cycle Thermodynamics
Equivalent Carnot Cycle
η = 1 - (Trej / Tadd)
Temperature
6
3
5
Tadd = Effective Tempe rature of Heat Addition
θ
= Log Mean Temperature of 2 & 3
= (T3 – T2) / ln (T3/T2)
Tadd
4
Trej = Effective Te mpe rature of Heat Rejection
= Log Mean Temperature of 4 & 1
3
2
1
= (T4 – T1) / ln (T4/T1)
4
2
Trej
1
s
Figure 4. Ideal, air-standard Brayton Cycle and its equivalent Carnot Cycle
Furthermore, we see that if Q units of energy, essentially equivalent to Q units of exergy were available
from the fuel (which is essentially equivalent to an infinite-temperature heat source), the exergy lost in the
heat addition process is Q Ta/Tadd and that rejected (but potentially recoverable) from the exhaust is Q (1Ta / Trej). Since the cycle's internal processes are all reversible, there is no internal exergy destruction and
the net work is the difference between the exergy added to the cycle and that rejected from it, i.e.
WNET = Q Ta (1 / Trej – 1 / Tadd )
.....…........ (18)
Noting that Ta = T1 and Q = m Cp (T3 – T2) and substituting from eqs. 15 and 16 for Tadd and Trej,
equation 18 then reduces to eq. 5.
2.2.3.3 Comparing the First & Second Law Approaches
The first law approach of §2.2.2 above stated that we had a certain amount of high-temperature energy
available (from fuel), we added it all to the working substance, converted some of it to work and rejected
the rest.
The second law approach of §2.2.3 gives the same answer but with a more insightful explanation. It
stated that we had equal amounts of energy and exergy available (from fuel), we added all the energy but
only part of the exergy to the working substance. A considerable portion of the exergy was lost in that
process due to degrading the heat down to Tadd. We then extracted part of the remaining exergy and
rejected the rest. The most important message of the Second Law approach is that major losses occur in
the "top end" of the cycle, i.e. heat addition, which appears 100 % energy efficient from a First Law view.
No ingenuity in designing the "bottom end" can recover those losses.
2-7
 M. A. Elmasri, 1990-2007
Chapter 2: GT Cycle Thermodynamics
2.2.4
EFFECT OF CYCLE PARAMETERS
2.2.4.1 Effect of Cycle Pressure Ratio
Raising Pressure Ratio:
Equivalent Carnot Cycle
η = 1 - (Trej / Tadd)
Raises T2 to T2' , which raises Tadd
Lowers T4 to T4', which lowers Trej.
Temp.
6
3'
The ratio Trej/Tadd is reduced.
η increases
3"
3
Turbine work 3'-4' > 3-4.
Compressor work 1-2' > 1-2.
Net Work may increase or decrease,
is maximum at an intermediate PR
5
θ
4
Tadd
2'
4"
3
4
2
2
1
2"
Trej
High η but
WC →WT
so Wsp → 0
4'
Low η
WC →0
WT→0
so Wsp→0
1
s
s
Figure 5. Effect of Brayton Cycle pressure ratio on heat addition & rejection temperature
Figure 5 illustrates how Brayton Cycle pressure ratio affects its heat addition and heat rejection
temperatures, hence its efficiency. Cycle 1-2-3-4 is the base case. Cycle 1’-2’-3’-4’ has a higher pressure
ratio than the base case, leading to a higher equivalent heat addition temperature and a lower equivalent
heat rejection temperature. Hence, it is more efficient. By contrast, Cycle 1’’-2’’-3’’-4’’ is less efficient
than the base case because it has a lower pressure ratio, leading to a lower heat addition temperature and a
higher heat rejection temperature. This provides additional insight into why ideal cycle efficiency
increases with pressure ratio.
Fig. 5 can also be used to understand the effect of pressure ratio on specific power. The three ideal cycles
shown have the same θ and a high, intermediate and low pressure ratio respectively.
With a high pressure ratio, 1-2’-3’-4’ the turbine and compressor work are both large but the difference
between them is smaller than at intermediate pressure ratios. Hence specific power suffers. As a further
illustration, the cycle diagram to the lower right of Fig. 5 shows an absurd extreme of a very high pressure
ratio, at which the compressor and turbine work approach equality, so their difference, the specific power,
approaches zero.
With a low pressure ratio, 1-2’’-3’’-4’’ both turbine work and compressor work are low, so their
difference is also low, resulting in low specific power. The cycle diagram to the lower right of Fig. 5
shows an absurd extreme of very low pressure ratio to further illustrate.
2-8
 M. A. Elmasri, 1990-2007
Chapter 2: GT Cycle Thermodynamics
The intermediate pressure ratio case, 1-2-3-4 has the maximum specific power.
Thus, one may conclude that raising the pressure ratio to obtain higher efficiency is a valid approach, but
an excessively high pressure ratio sacrifices specific power, thereby partially offsetting the benefit of
higher efficiency.
2.2.4.2 Effect of Cycle Maximum Temperature (T3)
Equivalent Carnot Cycle
η = 1 - (Trej / Tadd)
3'
T
3
Raising T3 to T3' :
Raises T4 to T4' , Tadd to Tadd' , and Trej to
Trej' , all in the same proportion.
Tadd'
The ratio Trej'/Tadd' = Trej/Tadd
η does not change.
Tadd
4'
2
Trej
Trej'
Turbine work 3'-4' > 3-4.
Compressor work 1-2 is same.
Net Work increases
4
1
s
Figure 6. Effect of Brayton Cycle maximum temperature on heat addition & rejection temperatures
Equation 7 showed that the thermal efficiency of the ideal cycle depends on pressure ratio only. That can
be appreciated by reference to Fig. 6 that depicts ideal cycles with low and high T3 but the same pressure
ratio. From a second law perspective, we observe that the log mean temperatures of heat addition and
heat rejection increase in the same proportion so that the ratio Tadd/Trej , hence the efficiency, remains
unchanged.
Another way to appreciate this result is to observe that raising θ (T3 ) also raises T4 , so both heat added
and heat rejected increase proportionately. The difference between heat added and rejected, which is the
net work output, increases proportionately as well. Hence specific power increases with higher values of
θ, but efficiency is unchanged.
2.3
REAL CYCLE ANALYSIS
To calculate through an actual gas turbine, the following effects that have been ignored in the ideal model
must be considered:
(i) Actual air, fuel and combustion gas properties.
(ii) Pressure drop in the combustor; as well as various, minor heat losses, mechanical losses, and
electrical losses from the gas turbine and generator.
(iii) Inefficiency in compressor and turbine due to fluid viscous friction.
2-9
 M. A. Elmasri, 1990-2007
Chapter 2: GT Cycle Thermodynamics
(iv) Extraction of cooling air from the compressor to cool the hot turbine components, and the
various losses associated with mixing the spent cooling air with the hot turbine gases.
2.3.1
DIFFERENCES BETWEEN IDEAL & REAL CYCLES
We shall first discuss the differences between the ideal analysis and the actual gas turbine listed in (i)-(iii)
above. This is done through an example of a cycle with a pressure ratio of 18 and a T3 of 1350 ºC for
which the calculated results are shown in Fig. 7 below.
In this example, the ideal air-standard cycle results show an efficiency of 56.2% and specific power of
545 kJ/kg, illustrated by the leftmost pair of bars in Fig. 7.
Figure 7. Differences between ideal and real cycle performance
Effect of real gas properties:
The second pair of bars in Figure 7 show the results if the real properties of air and combustion gases
were used instead of assuming constant air properties. It is still assumed that all components are ideal,
perfectly efficient and free of any pressure loss or friction losses. The result is a higher specific output
(662 kJ/kg, an increase of 22%) but a lower efficiency (a decrease of 4.6 percentage points, or 8%) than
the air-standard analysis.
The higher specific power arises because:
(a) When fuel is burnt instead of “adding heat”, the resulting hot flue gases have a higher specific heat
than air, so for a given temperature rise or drop, the gases in the turbine will have a greater enthalpy
difference than the air in the compressor.
(b) When fuel is burnt instead of “adding heat” the mass flow rate of combustion products going through
the turbine exceeds the mass flow rate of air going through the compressor (by roughly 2%).
The lower efficiency results from the same reasons. A larger heat input is needed to produce highenthalpy combustion products at T3 than to heat pure air to T3. Additionally, the enthalpy of the
combustion products leaving the exhaust at T4 exceeds that of pure air at the same temperature.
Effect of miscellaneous minor losses:
The third pair of bars shows the results after introducing all miscellaneous losses except that the
compressor and turbine are still assumed 100% efficient and without need for turbine cooling. These
2-10
 M. A. Elmasri, 1990-2007
Chapter 2: GT Cycle Thermodynamics
miscellaneous losses reduce specific power by 3.4% (from 662 kJ/kg to 640 kJ/kg) and reduce efficiency
by about 3.1% (from 51.6% to 50%).
Effect of turbomachine inefficiency:
Atmospheric
Air
Fuel
1
2
4
3
T
C
Exhaust
Gas
G
Isentropic
components:
WC = 40
WT = 100
WNET = 100 – 40 = 60
90% efficient
components:
WC = 40/0.9 ~ 44
WT = 100 x 0.9 = 90
WNET ~ 90 – 44 ~ 46
~24% less net work !
Figure 8. Illustrating the effect of turbomachine efficiency on real cycle output
The fourth pair of bars show the results after considering realistic compressor and turbine efficiencies.
Those have a large impact. Specific power drops by 21% (from 640 kJ/kg to 504 kJ/kg), and efficiency
drops by 17% (from 50% to 41.3%).
Fig. 8 helps to appreciate the impact of turbomachinery efficiency on cycle output. It shows a typical
example for which the isentropic (100% turbomachine efficiency) calculations give a turbine output of
100 units, the compressor absorbs 40 of them, leaving 60 for useful output. With 90% turbomachine
efficiency, the turbine makes 90 instead of 100 units of output, and the compressor absorbs 40/0.9 = 44.4
of them, leaving just 45.6 units for useful output. This represents a 24% reduction from the 60 units of
useful output corresponding to isentropic turbomachines.
Fig. 9 shows the performance of “real” gas turbine cycles, including all realistic effects except turbine
cooling. This is the performance to be expected if the turbine hot path components were made of
materials that could reliably withstand the temperatures shown without the need for cooling. Although
much research into such high-temperature, high-strength materials has been on-going, at the time of
writing these materials can still be described as “unobtanium”. However, it is the opinion of this author
that at some point in the future, materials developments, likely ceramics, will lead to practical gas
turbines that require little or no cooling.
2-11
 M. A. Elmasri, 1990-2007
Chapter 2: GT Cycle Thermodynamics
Brayton Cycle Gas Turbines with real effects except for assumption of a hypothetical
material that needs no cooling ("unobtanium" at present)
Cycle Efficiency @ Generator Terminals
48%
47%
46%
45%
44%
PR=12
PR=18
43%
PR=24
42%
PR=30
θ=4.592 (1050 °C)
41%
θ=5.112 (1200 °C)
40%
θ=5.633 (1350 °C)
39%
θ=6.154 (1500 °C)
38%
37%
36%
35%
150
200
250
300
350
400
450
500
550
600
650
700
Specific Power @ Generator Terminals kJ/kg (kW per kg/s)
Figure 9. Estimated performance of Brayton Cycle gas turbines with all realistic effects except turbine cooling
Comparing Fig. 9 and Fig. 3, we note that:
(i)
In the ideal air-standard cycle results of Fig. 3 the dashed lines for each pressure ratio are
horizontal, whereas in Fig. 9 they slope upwards. When turbomachinery losses are included,
efficiency tends to increase with higher T3 because this creates larger turbine work output for
a given compressor work. Thus, the turbine and compressor inefficiencies are less harmful
relative to the increased net output (specific power). Generally speaking, high specific power
diminishes the various component losses as a percentage of net output, which increases
efficiency as well as reduces sensitivity to component efficiencies.
(ii)
The specific power does not fall off with increasing pressure ratio in Fig. 9 by comparison
with Fig. 3. This is particularly true at higher values of T3. This is due to the fact that with
real gas properties, turbine gas enthalpies are much higher relative to compressor air
enthalpies, especially with increasing temperature.
Effect of turbine cooling:
The final pair of bars in Fig. 7 show results after considering turbine cooling. Specific power drops by
another 19% (from 504 to 406 kJ/kg) and efficiency by another 7% (from 41.3% to 38.6%). The turbine
cooling penalty is discussed below.
2.3.2
THE TURBINE COOLING PENALTY
At present, the turbine materials cannot sustain the stresses they are subject to with reliability and
longevity if they operate at temperatures above roughly 860 °C (1580 °F). Thus, they have to be cooled.
All older gas turbines, as well as many modern ones, use cooling air bled from the compressor and led
through the turbine blades to cool them internally. A schematic of a typical turbine blade with “internal
cooling” is shown in Fig. 10.
2-12
 M. A. Elmasri, 1990-2007
Chapter 2: GT Cycle Thermodynamics
Figure 10. Example of a cooled turbine blade (courtesy of GE)
In modern designs for the high-temperature stages, the spent cooling air is also injected from numerous
tiny holes along the blade surfaces to form a cooling film that helps reduce heat transfer from the hot
gases. Such “film cooling” allows a certain surface temperature to be achieved whilst using less cooling
air than “internal cooling” alone. Additionally, modern designs also incorporate a thin ceramic Thermal
Barrier Coating (“TBC”) on the blade surfaces to further economise on use of cooling air.
As pressure ratios have increased, resulting in higher compressor discharge temperatures and thus hotter
cooling air, several manufacturers have resorted to pre-cooling the hot, compressed cooling air, after
extracting it from the compressor and before routing it into the turbine blades. Whilst this reduces the
amount of cooling air needed, it increases the quenching effect on the hot gases flowing through the
turbine. Fig. 11 illustrates the various cooling options used in modern gas turbines.
2-13
 M. A. Elmasri, 1990-2007
Chapter 2: GT Cycle Thermodynamics
Figure 11. Cooling options in modern gas turbines
Naturally, a higher combustor exit temperature requires a higher cooling flow, which increases all losses
associated with turbine cooling, and these increased losses partially offset the advantage of the higher T3.
This result is illustrated in Fig. 12.
GT cooling effect on real Brayton Cycle, PR=18
GT cycle efficiency @ gen terminals %
42
41
Hypothetical no CA
CA uncooled
40
CA cooled H2O inj
CA cooled in HX
39
38
37
36
900
1000
1100
1200
1300
1400
1500
1600
Combustor Exit Temperature °C
Figure 12. Example illustrating the typical “cooling penalty” in modern gas turbines
Fig. 12 illustrates the impact of cooling on GT performance for a cycle with a PR=18 and designs with
increasing T3. The top line represents performance of a hypothetical, adiabatic machine that needs no
cooling, such as may one day be made of ceramics. The line labeled "CA uncooled" represents classical
GT technology, where the cooling air extracted from the compressor is used to cool the turbine directly, at
its extraction temperature. The line labeled "CA cooled H2O inj" is for designs where the cooling air
extracted from the compressor is cooled by water injection, before it is directed to the hot turbine
components. This results in a reduction in the amount of air needed to cool the turbine. However, that
reduced amount of coolant has roughly the same effect on the quenching of the hot gas path when it
mixes with the turbine gases, because although the amount is less, it is cooler. This tradeoff results in a
weak improvement in overall design performance at high values of T3, but no advantage at modest values
of T3. The line labeled "CA cooled in HX" is for designs where the cooling air extracted from the
compressor is cooled in a heat exchanger before it is directed to the hot turbine components. The heat
removed from the air is assumed lost in the results shown in Fig. 12. If the gas turbine were part of a
combined cycle, this heat can be utilised to generate steam, as in the combined cycle results shown in Fig.
13. If the gas turbine is not part of a combined cycle, this heat may be used to preheat the fuel, which
improves the apparent gas turbine efficiency by about 1/3 of a percentage point. The word “apparent” is
used in the previous sentence, because with warm fuel, the GT still consumes the same heat input, but a
small portion of this heat input is in the form of fuel sensible energy, so slightly less fuel is required.
One may see from Fig. 12 that if the sole objective of the design were maximum gas turbine cycle
efficiency, then the optimum combustor exit temperature would be in the 1250-1350 ºC (2280-2460 ºF)
range. Although higher temperatures create a cooling penalty that detracts from gas turbine cycle
efficiency, they are still desirable because:
2-14
 M. A. Elmasri, 1990-2007
Chapter 2: GT Cycle Thermodynamics
a) Higher temperatures give significantly greater specific power with only a modest sacrifice in gas
turbine cycle efficiency. The reduced capital cost in $/kW outweighs the reduction in efficiency in
most practical scenarios.
b) A higher T3 result in a higher T4 , which, when the gas turbine is used in a combined cycle, greatly
improves the efficiency of the steam bottoming cycle and hence of the combined cycle as a whole.
This is shown by Fig. 13.
GT cooling effect on combined cycle plant net electric efficiency
GT PR=18 with 3PRH bottoming cycle with typical design assumptions
63
62
Hypothetical no CA
CA uncooled
CA cooled H2O inj
CC net efficiency, %
61
CA cooled in HX
60
59
58
57
56
55
1200
1250
1300
1350
1400
1450
1500
1550
1600
Combustor Exit Temperature °C
Figure 13. Example illustrating the typical gas turbine “cooling penalty” on performance of modern combined cycles
Fig. 13 compares the performance in combined cycle of the engines whose simple-cycle performance is
shown in Fig. 12. The combined cycle efficiency continues to improve as T3 is raised beyond its value
for maximum efficiency of a simple-cycle GT. That is because the higher exhaust temperature provides
more recoverable energy for the steam cycle.
Gas turbines designed with a cooling air pre-cooler appear to offer a slight advantage in combined cycles
as seen in Fig. 13. In these calculations, the heat recovered from the cooling air is used to generate steam
to augment the bottoming cycle. The use of water injection to pre-cool the cooling air results in the
lowest combined cycle efficiency. Fig. 13 shows that a (hypothetical) adiabatic gas turbine can result in
combined cycle efficiency in the range of 62-63%.
2.3.3
PRINCIPAL ASSUMPTIONS FOR RESULTS WITH TURBINE COOLING
A word of caution is in order about the results of all calculations with turbine cooling presented here:
Calculations modelling cooled turbines require a complete, aero-thermodynamic design of an engine.
The results presented here were obtained with the THERMOFLEX software, which incorporates the
models of the author’s GASCAN program described in Ref. [1]. The inputs to this model include an
array of assumptions about material temperatures, thermal barrier coatings, film cooling effectiveness,
etc. These are a function of evolving technology, and the assumptions used here are the author’s
estimates of technology as of the time of writing (2007). They are by no means “accurate” in the sense of
modelling any particular manufacturer’s practice. For those interested in the assumptions used, the salient
ones are summarised below:
2-15
 M. A. Elmasri, 1990-2007
Chapter 2: GT Cycle Thermodynamics
Metal temperatures, stators
1st stage = 860 ºC
2nd and subsequent stages = 830 ºC
Metal temperatures, rotors
1st stage = 830 ºC
2nd and subsequent stages = 815 ºC
Uncooled stage efficiency
1st stage = 90%
2nd and subsequent stages = 90.5%
All cooled stages have film cooling and thermal barrier coatings
All stages have allowance for cooling endwalls, purging wheelspaces, and minor cooling air leakage
Cycles with PR=12 & 18 have 3-stage turbines, cycles with PR=24 & 30 have 4-stage turbines
Compressor polytropic efficiency = 91% - a conservative value that presumes some degradation
By way of example, Fig. 14 shows the calculated cooling air flow rates add up to about 18% of the
compressor inlet air for the case with a pressure ratio of 18 and a combustor exit temperature of 1500 ºC.
Additionally, the results shown for combined cycles depend on a vast array of assumptions. These results
were calculated with GT PRO, with most of its standard defaults for high-efficiency three pressure
reheat combined cycles with open loop water cooling at ISO conditions. A few inputs were changed from
their default values, chief amongst them is steam turbine efficiency which was increased from GT PRO’s
conservative default estimates.
Figure 14. Example of air distribution in a cooled gas turbine
2.3.4
TURBINE INLET TEMPERATURE DEFINITIONS
So far we have been using the notation T3 to imply, loosely, the combustor exit temperature and its
associated turbine inlet temperature. Indeed, the two are one and the same in the absence of turbine
cooling. With turbine cooling, additional definitions are needed.
Combustor exit temperature: This is the massflow-averaged total (stagnation) temperature of the gases
entering the first turbine stator row, after mixing any air used to cool the combustor transition pieces. For
an uncooled first-stator, such as for an old, low-temperature machine, or an advanced (hypothetical) one
with an uncooled ceramic first-row stator, this is the same as the total (stagnation) temperature entering
the first rotor. It is commonly called “Turbine Inlet Temperature” and abbreviated as “TIT”. However,
it is not unusual for the abbreviation “TIT” to denote the turbine rotor inlet temperature, defined below.
In the current work, we shall reserve the abbreviation “TIT” to imply combustor exit temperature.
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 M. A. Elmasri, 1990-2007
Chapter 2: GT Cycle Thermodynamics
Turbine Rotor Inlet Temperature (T3): Since T3 in cycle thermodynamic analysis is the temperature at the
beginning of turbine work-extraction, it is the stagnation temperature entering the first rotor. This is
usually called "Firing Temperature". It is also sometimes spelled out in full and called “Turbine Rotor
Inlet Temperature” and abbreviated as “TRIT”. Its not uncommon, however, to refer to it by the
abbreviation “TIT”. In the current work, we shall refer to it by the abbreviations “T3” or “TRIT”.
With cooled turbines, the combustor discharge mixes with the first stator cooling air, increasing in
massflow and decreasing in temperature, before impinging upon the first rotor. For typical, modern
engines, 4-7% of the compressor inlet airflow is used to cool the first stator row and first wheelspace,
resulting in T3 below combustor exit by some 40-70 ºC (72-112 ºF).
The first-stator and first wheelspace cooling air is indistinguishable thermodynamically from the
combustor dilution air and transition piece cooling air. Both mix with the combustion products resulting
in the total (stagnation) temperature T3 entering the first rotor, which begins the work extraction process.
Thus, for a given value of T3, cooling the first stator and wheelspace can be considered immaterial to the
cycle, and the cooling air used in the first stator and first wheelspace is called "non-chargeable cooling
air”. This term is only strictly correct if the total pressure loss engendered by mixing the stator and first
wheelspace coolant with the high-speed gases is added to the combustor pressure drop. “Non-chargeable
cooling air” should still be minimised in an efficient design to reduce the gas stagnation pressure loss
incurred by mixing slow-moving coolant with the high-speed gas. Also, if more air is used to cool the
first stator, a higher combustor exit temperature is needed to achieve a given T3, exacerbating the problem
of cooling the combustor liner and transition pieces. In conclusion, “non-chargeable cooling air” still has
some costs.
ISO Inlet Temperature (Tiso): T3 is not a perfect measure of cycle working temperature, since for a given
T3, if a great deal of cooling air is used in the first rotor and subsequent stages, performance will still
suffer from the excessive quenching of the hot working gases. A definition that attempts to represent this
quenching in a reasonable fashion is the "ISO Turbine Inlet Temperature” described in German
Standards DIN 4341. This is the temperature that would result from mixing all cooling air streams into
the combustor discharge and is 60-120 ºC (108-216 ºF) below T3 for typical, modern engines.
Typical differences between the turbine inlet temperature definitions, based on a
cycle with a PR=18
0
Tiso
Difference from combustor exit, °C
-20
T3 (TRIT)
-40
-60
-80
-100
-120
-140
-160
-180
-200
900
1000
1100
1200
1300
1400
1500
1600
Combustor Exit Temperature °C
Figure 15. Typical differences between turbine inlet temperature definitions
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 M. A. Elmasri, 1990-2007
Chapter 2: GT Cycle Thermodynamics
Some manufacturers do not publish their turbine inlet temperatures. Those who do may use different
definitions. US-based manufacturers, as well as some others, tend to use “Firing temperature”, meaning
T3 as the standard definition. German and Swiss manufacturers tend to use Tiso. Japanese manufacturers
tend to use TIT (combustor exit). To further add to the confusion, some use terms like “Turbine Inlet
Temperature” or abbreviations like “TIT”, or terms like “Firing Temperature”, to mean any of the above.
It does not matter which definition is used in any qualitative discussion, since all increase hand-in-hand.
In a quantitative discussion or calculation, however, it is imperative to be unambiguous.
Naturally, in the absence of turbine cooling, the three definitions given above would be identical. With
turbine cooling, the differences are illustrated by Fig. 15 for the example cycle with a pressure ratio of 18.
Naturally, the difference between the definitions increases with higher TIT since that difference derives
from the cooling air flow, which increases with hotter gases. Tiso deviates more steeply from TIT as TIT
increases because Tiso reflects the coolant flow to additional stages beyond the first stator, whereas T3
only reflects the increasing burden of cooling the first stator alone.
2.3.5
RESULTS FOR THE REAL CYCLE MODEL
Real Brayton Cycle Gas Turbines with Direct Air Cooling
Cycle Efficiency @ Generator Terminals
42%
41%
40%
PR=12
39%
PR=18
PR=24
38%
PR=30
θ=4.592 (1050 °C)
37%
θ=5.112 (1200 °C)
θ=5.633 (1350 °C)
36%
θ=6.154 (1500 °C)
35%
34%
33%
150
200
250
300
350
400
450
500
Specific Power @ Generator Terminals kJ/kg (kW per kg/s)
Figure 16. Estimated performance of Brayton Cycle gas turbines with all realistic effects including turbine cooling
Fig. 16 shows the efficiency vs. specific power plot for the real cycle model. As the design TIT is
increased, the efficiency peaks then begins to fall as the cooling penalty offsets the advantage of the
increased specific power. Please contrast Fig. 16 with Fig. 9 to see the effect of turbine cooling on gas
turbine cycles.
Increasing the design TIT (and cooling flow) beyond the maximum efficiency point continues to increases
specific power, until, at some point beyond the limit of the curves of Fig. 16, the specific power would
peak then begin to fall as well. Generally, most modern engines are designed to operate at TIT's beyond
those for peak efficiency in order to raise the specific power and/or exhaust temperature for applications
with heat recovery. Design total cooling air flows are typically about 10-20 % of compressor inlet
airflow, including the "non-chargeable" air to cool the first stator, which is typically 4-7% of inlet airflow.
Thus, “chargeable cooling air” is typically 6-13% of compressor inlet airflow. Improvements in materials
and blade cooling technology are continuously reducing the need for cooling air and raising the design
TIT.
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 M. A. Elmasri, 1990-2007
Chapter 2: GT Cycle Thermodynamics
2.4
STEAM COOLED GAS TURBINES FOR COMBINED CYCLES
Fig. 13 shows that cooling the GT costs 3-4 percentage points of combined cycle efficiency. One method
of reducing that loss is to cool the GT internally using steam in a closed loop, with the energy extracted
from the turbine gas path being utilised in the steam portion of the combined cycle.
Internal steam cooling of the turbine blades is feasible only with a ceramic thermal barrier coating (TBC)
applied to the exterior of the blade surfaces. These coatings reduce the heat flux through the blades,
which, in the absence of these insulating coatings, would result in excessive temperature gradients and
thermal stresses in the blades. TBC technology is well proven and widely applied in gas turbines.
However, when a TBC is used in conjunction with film-cooling or modest gas temperature, it is basically
a life-augmentation measure, which extends blade life by reducing metal temperature and shielding the
metal from potential corrosion/erosion. In this augmentation role, gradual loss of the TBC by erosion or
spalling does not lead to blade failure. When a TBC is used with very hot gases and with blades that are
solely cooled internally, and lacking the heat flux reduction associated with film-cooling, the TBC
becomes a vital component rather than an augmenting measure. In such a critical use, reliability and
monitoring of the TBC condition are necessary.
Figure 17a. THERMOFLEX model of a steam-cooled gas turbine integrated with a GT PRO model of a steam
bottoming cycle shown in Fig. 17b.
The benefits of steam cooling are:
1. Increased specific power, since most of the compressor inlet airflow passes through the initial turbine
stages when air-cooling is minimised.
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 M. A. Elmasri, 1990-2007
Chapter 2: GT Cycle Thermodynamics
2. Increased excess air through the combustion section, allowing better cooling of liner and transition
piece for a given TIT.
3. Increasing T3 for a given TIT, since the gas temperature drop as it flows through a TBC-insulated
steam-cooled first-stage nozzle is much less than with an air-cooled first-stage nozzle in which the
cooling air is mixed with the gases.
4. Higher T4 (turbine exhaust temperature) for a given T3. This helps to produce more steam in the
HRSG of a combined cycle.
5. Higher steam production in the HRSG of a combined cycle due to utilisation of the energy removed
from the gas path in the steam system.
6. Reduced gas path losses in stagnation pressure due to mixing of the cooling air with the gas flow. This
benefit is partly offset by the increase in aerodynamic losses due to the thick blade trailing edges
associated with an internally steam-cooled blade.
Figs. 17a & 17b show a THERMOFLEX/GT PRO model for a steam-cooled GT integrated with a
combined cycle. The GT has a pressure ratio of 23:1, with a four-stage turbine. The first two stages are
internally steam-cooled, except for wheelspaces, which are air-cooled. The third stage as well as the
wheelspaces of the fourth are air-cooled. The cooling steam originates from the cold reheat of the steam
portion of the combined cycle, and, after passing through the blades, is returned to the hot reheat steam
flowing to the steam turbine. Roughly 1/3rd of the hot reheat steam passes through the GT in this
calculation. About 5% of the cooling steam is assumed to leak into the GT gas path.
The modelling calculations shown in Figs. 17a & 17b produce results that roughly approximate those
given in several articles in the trade press, published in 2000, about the GE 9H combined cycle then under
construction at Baglan Bay, in Wales. Since very little detailed information was given in these articles,
the assumptions used in the models of Figs. 17a & 17b are based solely on judgment and estimates of the
present writer, rather than a reflection of the actual GE design.
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 M. A. Elmasri, 1990-2007
Chapter 2: GT Cycle Thermodynamics
Figure 17b. GT PRO portion of the THERMOFLEX model of a steam-cooled gas turbine combined cycle
Figs. 18 through 20 illustrate the benefits of steam cooling relative to air-cooling a Brayton Cycle GT
with a pressure ratio of 23:1, a TIT of 2630 °F (1443 °C), and a 4-stage turbine. Five sets of bars are
shown on each of these figures. The first set of bars applies to the base case, where all blade rows are aircooled. The second set applies to a design where only the first stage stator blades are steam cooled,
instead of being air-cooled, but with all other blade rows air-cooled. The third set of bars applies to a
design where the first stage stator and rotor are both steam-cooled. The fourth applies to a design where
steam-cooling is extended to the second-stage stator, and the fifth to a design where the steam cooling
extends to the second-stage rotor as well.
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 M. A. Elmasri, 1990-2007
Chapter 2: GT Cycle Thermodynamics
Figure 18. Effect of steam-cooling the gas turbine on combined cycle efficiency
Fig. 18 shows the increase in combined cycle efficiency as steam-cooling is used instead of air-cooling
for an increasing number of blade rows. Two bars are shown in each case, one for designs where the
cooling air is used directly, and one for the designs where the cooling air is pre-cooled in an EPC. Since
these results are obtained for a fixed TIT, steam-cooling the first stator results in an increase in T3 (1st
rotor inlet temperature), which improves performance by a greater amount relative to air cooling in the
EPC cases, since the increased T3 tends to increase 1st rotor's and subsequent blade rows' cooling
requirements. Steam-cooling the first two stages improves the combined cycle efficiency by about 3
percentage points relative to air-cooling.
Figures 19 and 20 provide further insight into the results of Fig. 18, the former for direct air-cooling and
the latter for air-cooling with EPC.
Fig. 19 shows that steam-cooling the first two stages increases GT specific power by 25%, relative to
direct air-cooling. It increases ST specific power per unit of GT inlet airflow by 10%, due to the higher
exhaust temperature from the GT as well as the energy added to the reheat steam in the steam portion of
the combined cycle. The net result is a 20% increase in combined cycle specific output.
Fig. 20 shows that steam-cooling the first two stages increases GT specific power by only 15%, relative to
air-cooling with EPC, since the reduction of water injection associated with a reduction in cooling airflow
diminishes the specific power gained by steam-cooling. It increases ST specific power per unit of GT
inlet airflow by 8%, rather than 10% for direct air-cooling; and combined cycle specific power by 13%,
rather than 20% for direct air-cooling.
2-22
 M. A. Elmasri, 1990-2007
Chapter 2: GT Cycle Thermodynamics
Figure 19. Power increase as more GT blade rows are steam-cooled, instead of air-cooled. The air-cooled blades
use air directly from the compressor bleeds.
Figure 20. Power increase as more GT blade rows are steam-cooled, instead of air-cooled. The air-cooled blades
use air from the compressor after precooling it by water injection
•
High specific power, due to a greater percentage of the airflow throuh the 1st stage
•
More dilution air available for combustor liner & transition piece cooling
•
Lower TIT (Combustor discharge) for the same T3 (1st Rotor Inlet)
•
Higher exhaust temperature due to less quenching of the gases with cooling air
•
Reduced pressure losses in the hot gas due to mixing with spent cooling air
•
Improves steam production in the HRSG by heating part of the reheat steam
Fig. 21 summarises the advantages of steam-cooling the GT.
2-23
 M. A. Elmasri, 1990-2007
Chapter 2: GT Cycle Thermodynamics
2.5
THE REHEAT GAS TURBINE CYCLE
Reheating the combustion gases as they expand through the turbine raises the mean cycle heat addition
temperature and improves the fundamental thermodynamics.
Reheat may be done semi-continuously. An experimental turbine with hydrogen cooling of the stators and
its subsequent combustion in their wakes has been tested [26]. This takes the most advantage of the peak
cycle temperatures and approaches Carnot efficiency [27,28]. It enables high efficiency with moderate
temperatures, requiring little or no cooling. If ceramic structures become a practical reality very high
efficiencies could be reached. A single reheat by one additional combustor requires less deviation from
existing design practice. A large (120 MW), intercooled, high-pressure-ratio (55:1), prototype machine
was tested in the mid-1980's [29-32]. Analysis by the present writer, done in the early 1980's [33]
indicates that with the technology available or under development at that time, the increased cooling
losses diminish the efficiency advantage of reheat. This showed that the intercooled/reheat gas turbine
gives combined cycle efficiencies similar to Brayton Cycle gas turbines, but produces 60% more specific
power in combined cycle. The same analysis showed that the non-intercooled, reheat GT promised
higher efficiency combined cycles than Intercooled/Reheat or Brayton Cycle GT’s.
The non-intercooled reheat cycle GT was brought to commercial reality in the 1990’s by ABB (now
Alstom). ABB called it "Sequential Combustion", rather than "Reheat", presumably to avoid confusing
the Reheat GT Cycle with reheat in the steam portion of the combined cycle. A schematic and a photo of
this machine is shown are Figures 22a & 22b.
Fig. 22a Cutaway view of GT24/GT26 (Courtesy of Alstom – Formerly ABB)
2-24
 M. A. Elmasri, 1990-2007
Chapter 2: GT Cycle Thermodynamics
Fig. 22b. Schematic sectional view of GT24/GT26 & photo of the machine (Courtesy of Alstom – Formerly ABB)
To compare the intrinsic merit of the Reheat GT Cycle with the Brayton Cycle, one needs to place both
cycles on an equal footing with respect to component technology, since, obviously, a cycle which is
intrinsically inferior in thermodynamic design may appear to be superior solely because it is implemented
with more efficient turbomachines, or higher blade temperatures, or better blade-cooling design, etc. The
results presented here were obtained with the GASCAN cooled-turbine model (Ref. [1]) now incorporated
into the THERMOFLEX software. The model requires numerous inputs to characterize the technology,
such as maximum allowable metal temperature for each blade row, the cooling effectiveness curve for
each blade row, the compressor polytropic efficiency, the adiabatic efficiency of each turbine stage, the
number and location of cooling air bleeds, etc. By using the same inputs for different cycles, one may
compare the intrinsic relative merits of the thermodynamic cycles themselves, without specific knowledge
or reference to any particular manufacturer’s practices, and without biasing the results by interjecting
non-identical component technologies into a discussion of thermodynamic cycles.
Figure 23 shows a schematic of the THERMOFLEX model of a reheat GT with PR=32 and TIT=1320 ºC,
and in which the turbine cooling air extracted from the compressor at the two highest-pressure bleeds is
cooled in heat exchangers before being utilised for turbine cooling.
Fig. 23. THERMOFLEX model of a reheat GT with PR=32 and TIT=1320 ºC. The model is generic and does not
accurately represent the Alstom technology of the GT24/26, despite having the same pressure ratio
Fig 24 shows the h-s cycle diagram corresponding to the cycle model of Fig. 23. Because the enthalpy
datum used is zero at 25 ºC, the air inlet at 15 ºC (state 1) shows negative specific enthalpy. The
compression line (1-2) has marks to indicate cooling air bleeds. The primary combustor (2-3a) produces
hot gases at state 3a., which represents the combustor exit (not rotor inlet) state. The drop in total
pressure and temperature due to cooling the stator between combustor exit and rotor inlet is calculated
internally within the model. The single stage HP turbine expands the gases to extract power, whilst
2-25
 M. A. Elmasri, 1990-2007
Chapter 2: GT Cycle Thermodynamics
mixing the HP turbine’s cooling air into the gases. State 3b leaving the HP turbine has a lower specific
entropy than state 3a, despite the considerable entropy generation in the HP turbine, because mixing low
temperature (low specific entropy) coolant outweighs the entropy generation. The specific enthalpy drop
(3a-3b) does not directly equal turbine specific work, because part of that drop arises from mixing lowenthalpy coolant; besides, the mass flow rate at state 3b includes the coolant, whereas the mass flow rate
at state 3a excludes it. The secondary (reheat) combustor (3b-3c) produces hot gases at state 3c which
then enter the 4-stage LP turbine. State 3c represents secondary combustor exit, not rotor inlet to the next
turbine stage. The LP turbine expansion line (3c-4) has marks to indicate states between its stages.
h-s diagram of a reheat GT cycle with PR=32 & TIT=1350 C
1800
3a
Compressor
1600
3c
Combustor
HP turbine
1400
Reheat Combustor
Turbine
Heat rejection
Specific enthalpy, kJ/kg
1200
3b
1000
800
2
600
4
400
200
0
1
-200
0
0.2
0.4
0.6
0.8
1
1.2
1.4
1.6
Specific entropy kJ/kg-K
Fig. 24. h-s diagram of the reheat GT cycle shown in Fig. 23
Figure 25 shows the effect of cooling on the reheat GT cycle, and it should be compared with Fig. 12, its
equivalent for Brayton cycle GT’s. In the hypothetical case of no turbine cooling, the intrinsic superiority
of the reheat cycle is about 3 percentage points (~ 44½ % vs. ~ 41½ % with the same set of component
assumptions). This is due to its heat addition processes (2-3a and 3b-3c) occurring at a higher mean
temperature than the typical process (2-3) of a Brayton cycle.
However, the reheat cycle, with its two “hot stages”, as well as its higher pressure ratio, must suffer a
greater cooling penalty than the Brayton cycle. High pressure ratio requires more turbine stages that need
cooling, and results in hotter cooling air at the compressor exit. If the hot cooling air is used directly, an
excessive amount is needed. If the hot cooling air is pre-cooled, less is needed, but the pre-cooling
process involves losses and destroys exergy. Thus, comparing Figs. 12 and 25, we see that the cooling
penalty for the reheat cycle GT is about 6 percentage points, vis a vis 3 percentage points for the Brayton
cycle GT. However, the reheat cycle’s 6-point cooling penalty is applied to a cycle that has an intrinsic
thermodynamic superiority of 3 points, so after cooling penalties, both the reheat and Brayton cycle GT’s
end up with about the same efficiency when compared on an equal footing, with the same component
assumptions.
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 M. A. Elmasri, 1990-2007
Chapter 2: GT Cycle Thermodynamics
GT cooling effect on RHGT Cycle, PR=32
45
GT cycle efficiency @ gen terminals %
44
43
Hypothetical no CA
42
CA uncooled
CA cooled H2O inj
41
CA cooled in HX
40
39
38
37
36
900
1000
1100
1200
1300
1400
1500
1600
Combustor Exit Temperature °C
Fig. 25. Example illustrating “cooling penalty” in a reheat gas turbine. Please compare with Brayton cycle gas with
the same assumptions, as shown in Fig. 12.
The fact that the reheat cycle GT and the Brayton cycle GT have about the same efficiency should not be
misinterpreted to construe that they both allow the same combined cycle efficiency.
Efficiency parity between two gas turbines implies that both have the same heat rejection as a percentage
of fuel heat input. If they differ in specific power, the one with a greater specific power will, necessarily,
also have a higher exhaust temperature. This follows from the fact that greater specific power with the
same efficiency means more power produced per unit airflow, proportionately more fuel input per unit
airflow, and proportionately more exhaust energy per unit airflow. Thus, the GT with higher specific
power will have a hotter exhaust, and a greater percentage of its exhaust heat rejection can be more
efficiently recovered in a HRSG of a bottoming cycle.
Thus, if one compares two gas turbines with the same efficiency but different specific power, then, by
definition, the one with higher specific power will have a higher exhaust temperature, a higher exergy
efficiency, and will allow more effective heat recovery and a more efficient combined cycle.
High GT specific power also provides a specific cost advantage in combined cycle because it results in a
larger percentage of the combined cycle’s output emanating from the gas turbine and a smaller percentage
emanating from the steam turbine. Generally, in a combined cycle, the gas turbine cost is about 1/3rd of
the total and it provides 2/3rds of the output. This implies that capital cost of the GT per unit of its own
power output is about ¼ of the capital cost of the bottoming cycle per unit of its steam turbine’s output.
Fig. 26 shows that when compared with the same component assumptions, and with pressure ratios wellsuited to each type of cycle, a reheat GT cycle possesses an advantage of about 15% in specific power
over a Brayton cycle GT, and, by extension, the exhaust temperature of the reheat cycle GT is about 50 ºC
higher than of the Brayton cycle GT. This provides an advantage in a combined cycle.
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 M. A. Elmasri, 1990-2007
Chapter 2: GT Cycle Thermodynamics
Specific power & exhaust temperature comparison between RH cycle (PR=32)
and Brayton cycle (PR=18) GT's. Both have CA heat exchangers
Specific power, kJ/kg & Exhaust temp. °C
750
700
650
600
550
500
450
400
350
RH GT specific power
300
Brayton GT specific power
RH GT T4
250
Brayton GT T4
200
900
1000
1100
1200
1300
1400
1500
1600
Combustor Exit Temperature °C
Fig. 26. Example illustrating the specific power and exhaust temperature advantages of a reheat gas turbine when
compared with a Brayton cycle gas of similar technology
GT cooling effect on combined cycle plant net electric efficiency
Reheat GT PR=32 with 3PRH bottoming cycle with typical design assumptions
65
64
CC net efficiency, %
63
Hypothetical no CA
CA cooled in HX
62
61
60
59
58
57
56
1200
1250
1300
1350
1400
1450
1500
1550
1600
Combustor Exit Temperature °C
Fig. 27. Example illustrating combined cycle performance of reheat GT. Please compare to the equivalent chart of
Fig. 13 for similar combined cycles employing a Brayton cycle GT
Fig. 27 shows the combined cycle efficiency possible with a reheat GT. In the limit of a (hypothetical)
adiabatic gas turbine, combined cycle efficiency of 64-65% would be attained, about 2 percentage points
higher than the 62-63% of Brayton cycle GT’s with the same assumptions (Fig. 13) After considering
turbine cooling, with TIT in the 1450-1550 ºC range, the reheat GT combined cycle efficiency is in the
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 M. A. Elmasri, 1990-2007
Chapter 2: GT Cycle Thermodynamics
59½ - 60% range, vs. 58½ - 59% for the Brayton cycle GT, when both are calculated with the same
component assumptions. Thus, the reheat GT cycle offers an intrinsic efficiency advantage of about one
percentage point in a combined cycle, besides the specific power advantage of about 15%.
In Fig. 27, the curve showing results with turbine cooling is for the case of a cooling air pre-cooler
recovering heat from the cooling air and generating steam to augment the bottoming cycle. Thus, it
should be compared with the equivalent curve labelled “CA cooled in HX” in Fig. 13.
The table below summarises the performance comparison between the Reheat and Brayton cycles for the
case with TIT = 1450 ºC (2642 ºF). It should be again emphasized that these results are calculated by the
author and do not reflect any particular manufacturer’s products or technology.
RH GT
38.2%
511 kJ/kg (232 kW per lb/s)
670 °C (1240 °F)
59.5%
GT gross efficiency
GT specific power
GT exhaust temperature
CC net efficiency
2.6
Brayton GT
38.3%
448 kJ/kg (203 kW per lb/s)
625 °C (1157 °F)
58.5%
RECUPERATED AND INTERCOOLED GAS TURBINE CYCLES
A recuperate is a heat exchanger that preheats the compressor delivery air before the combustor, saving
fuel and improving efficiency. The heat source is the turbine exhaust gases. Fig. 28 illustrates the cycle
and its T-s diagram.
Cooled
Exhaust
Gas
4a
Atmospheric
Air
1
2
Temperature
2a
C
3
6
4
Fuel
5
Exhaust
Gas
3
T
G
Cooling Air
θ
4
4
2a
3
2
Available
ΔT
4a
2
1
1
s
Fig. 28. Recuperated GT cycle
Sometimes the recuperator is called "regenerator". The preferred nomenclature is recuperator for a fixed
surface heat exchanger and regenerator for a moving media device, which passes through the hot and cold
stream periodically, transferring heat from one to the other. In what follows, "recuperator" is used, but the
arguments apply just as well for a regenerator.
For recuperation to work, the turbine exit temperature T4 must be higher than that at compressor delivery
T2. This implies a pressure ratio below that for maximum specific power. For a given T3, lower pressure
ratios raise T4 and depress T2, increasing the available ∆T for heat recovery by the recuperator. This
increases the efficiency of an ideal recuperated cycle. For the real case, the reduction in specific power at
low pressure ratios increases the various losses as a proportion of net power. This is exacerbated by the
2-29
 M. A. Elmasri, 1990-2007
Chapter 2: GT Cycle Thermodynamics
pressure losses in the recuperator itself with the result that the efficiency advantage of this cycle is limited
to a narrow range of modest pressure ratios.
Recuperation and raising the pressure ratio have the same, fundamental thermodynamics effect on
efficiency, since both raise the mean heat addition temperature and depress the mean heat rejection
temperature. Both "recycle" energy from the back end of the turbine to back end of the compressor. The
recuperated cycle moves it as heat, requiring a heat exchanger of considerable area. The high-pressureratio cycle moves it as work via. a few additional blade rows of turbomachinery.
Cooling
Wate r
2'
Atmospheric
Air
4a
2"
LPC
4
Fuel
2
1
Temperature
Cooled
Exhaust
Gas
2a
3
T
HPC
Exhaust
Gas
G
3
6
Cooling Air
5
θ
4
4
2a
3
Available
ΔT
2
1
2
2"
4a
2'
1
s
Fig. 29. Intercooled/Recuperated GT cycle
Until the 1980’s, gas turbine pressure ratios were modest, and recuperated gas turbine cycles up to about
30 MW were common, particularly for mechanical drive gas turbines in gas pipelines, refineries, and
industry. Since the 1980's, the promulgation of high-pressure-ratio aeroderivatives has reduced the appeal
of the low-pressure-ratio recuperated cycle; except for small, vehicular gas turbines, which are physically
too small to have high pressure ratios. Currently, the non-intercooled recuperated cycle is applied in
practice in smaller gas turbines, from a few kW for micro-generation to a few MW for specialised vehicle
or marine propulsion. Its use in larger, industrial and pipeline mechanical drive applications, of a few
tens of MW, has been largely superceded by aeroderivatives.
To gain the efficiency advantage of recuperation, without sacrificing specific power by using a low
pressure ratio, the compressor can be intercooled, as shown in Fig. 29. Intercooling the compressor
lowers the total work of compression, which increases net cycle work. The cooler compressor discharge
does not require more fuel to be heated to T3, because it recoups more heat from the exhaust gases. This
provides a substantial boost to both specific power and efficiency.
Fig. 30 shows performance map of recuperated cycles, with and without intercooling. The component
assumptions used are from early 1990’s technology, and typically produce efficiency about one
percentage point below those used in §2.1 through §2.5. The non-intercooled recuperated cycle is best
with a pressure ratio of ~ 12, and with conservative assumptions gives a gross efficiency (at the generator)
of about 43% and specific power of about 350 kJ/kg. (160 kW per lb/s). This compares favourably with a
modern, high pressure ratio aeroderivative.
2-30
 M. A. Elmasri, 1990-2007
Chapter 2: GT Cycle Thermodynamics
Recuperators may come back into practice for mid-range machines in conjunction with intercooling,
water injection and/or steam injection. The Intercooled Recuperated (ICR) is attractive for mid-size
applications, such as ship propulsion [4-6]. It can attain 47%-48% efficiency with a specific power 30%
higher than for a Brayton Cycle.
220
Efficiency @ Gen. Term., %
48
265
309
353
397
441
485
529
kJ/kg
INTERCOOLED
RECUPERATED
(ICR)
46
TIT=2100-2700 F by 200
TIT=1700-2100 F by 100
2100-2500 F by 200
44
PR = 8
PR = 12
PR = 18
PR = 24
TIT=1900-2700 F by 200
42
RECUPERATED
40
1700 F =
1900 F =
2100 F =
2300 F =
2500 F =
2700 F =
TIT=2100-2700 F by 200
38
80
100
120
140
160
180
200
220
927 C
1038 C
1149 C
1260 C
1371 C
1482 C
240
Specific Power @ Gen. Term., kW per lb/s
Fig. 30. Recuperated GT cycle performance
One recent development by GE is the LMS100, an intercooled, but non-recuperated gas turbine with
high-pressure-ratio (~40:1). This provides an efficiency of about 45% and is depicted in Figure 31. With
its high efficiency and modest exhaust temperature, this machine is well-suited to cogeneration, district
hating, desalination, and feedwater repowering of conventional steam power plants.
Fig. 31. Intercooled, but non-recuperated GT (GE LMS100 courtesy of GE)
2-31
 M. A. Elmasri, 1990-2007
Chapter 2: GT Cycle Thermodynamics
2.7
WATER-INJECTED RECUPERATED (WIR) GAS TURBINE CYCLES
Refs. [7-11] discuss this cycle and some variations. Here water is injected after compression and again in
the recuperator. This cools the compressed air and extracts more heat from the exhaust gases, lowering
the final heat rejection temperature. The additional water vapor mass flow rate increases turbine power
by utilising the additional heat recovered at the low temperature end of the exhaust gas.
Another variation, known as the "HAT Cycle" (for Humid Air Turbine) is described in Ref. [13]. The
layout is more complicated than that of the WIR cycle, but its efficiency is essentially similar.
Various claims have been made that the HAT cycle is more efficient than a combined cycle. However,
the present author has calculated both cycle types using the same component technology, and reasonable
assumptions for each cycle’s unique parameters. The present author’s conclusion is that when compared
on an equal footing, a triple-pressure reheat combined cycle is about 3 percentage points more efficient,
and comparable in specific power to a WIR/HAT cycle. The combined cycle, however, is likely to be
more expensive than a commercially-mature WIR/HAT cycle. Thus, the WIR/HAT cycle may present an
economically attractive option in the range of 20-120 MW, where the efficiency advantage of the
combined cycle is reduced and its specific cost increased.
2.8
STEAM-INJECTED GAS TURBINE CYCLES
This class of GT Cycle recycles exhaust heat using steam as a medium. It is sometimes known as the
"Cheng Cycle", after Dr. Cheng who obtained several patents in this area, or as the "STIG Cycle", its
name trademarked by GE. Although a limited amount of steam, about 3% of inlet airflow, is frequently
injected into the GT combustor for NOx control, a Steam-Injected Cycle is defined as one where the
steam injection is primarily for thermodynamic performance, in the full amount that can be generated by
the hot exhaust gases, and which is much greater than needed just for NOx control. In practice, it needs
an engine specially designed or modified for it, in order to be able to ingest all the steam generated by the
exhaust heat. The Cheng/STIG cycle has gained a certain but limited measure of acceptance. Refs. [1425] show some of the developments and applications.
2.8.1
SINGLE-PRESSURE STEAM-INJECTION
Fig. 32. “Cheng” or “STIG” Cycle
Fig. 32 illustrates the system. The steam generated in a heat recovery boiler is injected into the gas
turbine cycle at combustor pressure. This increases the mass flow through the turbine, with a large
relative increase in specific power. Since the steam must be heated from the injection temperature to T3,
additional fuel is required, and the relative increase in efficiency is typically less than half the relative
increase in specific power. A crude (but useful) rule of thumb is that all other things remaining constant,
an increment of steam injection of 1% of inlet air massflow will increase power by 2% and decrease
apparent heat rate by 1%.
2-32
 M. A. Elmasri, 1990-2007
Chapter 2: GT Cycle Thermodynamics
Fig. 33. Heat recovery temperature profile in a single-pressure steam-injected gas turbine cycle
Whether the steam is mixed with the fuel or injected into the compressor discharge air or the combustor
makes no thermodynamic difference as long as T3 is the same. The location of steam injection, however,
influences the flame temperature, the formation of pollutants, the stability of combustion, and the cooling
of the combustor liner.
Maximum cycle performance is obtained if all the steam that could be generated by the exhaust gases
were injected into the machine. Generally, this would require redesigning an standard engine so that the
turbine's size would be greater relative to the compressor's.
Injection Steam Production Capability, % of Airflow
35
30
25
20
TIT = 2300 F (1260 C)
TIT = 2000 F (1093 C)
15
10
5
0
8
12
16
20
24
28
32
36
GT Pressure Ratio
Fig. 34. Steam generation capability for single-pressure steam-injected gas turbine cycles as a function of GT cycle
pressure ratio
Fig. 33 shows the heat recovery profile for the heat recovery boiler of Fig. 32. The large mean
temperature difference between the exhaust gas as it cools and the boiler steam side implies a high exergy
loss. This is manifested in the energy balance by the large amount of latent heat discharged at the stack.
The figure also illustrates the principal constraint on the thermodynamics of this cycle. To get more
benefit from a unit mass of steam injection into the turbine requires the highest possible GT pressure
ratio. This necessitates a higher pressure and saturation temperature in the heat recovery boiler, reducing
2-33
 M. A. Elmasri, 1990-2007
Chapter 2: GT Cycle Thermodynamics
the amount of steam that can be generated by the hot gas by diminishing the effective ∆T, as shown on
Fig. 33. Raising the pressure ratio of the GT also reduces its exhaust temperature, further diminishing
that effective ∆T. Fig. 34 shows how the exhaust steam generation capability of a steam-injected gas
turbine falls with increasing its pressure ratio.
Performance of the STIG cycle and its intercooled, version, sometimes referred to as ISTIG, is shown in
Fig. 35. The component assumptions used are from early 1990’s technology, and typically produce
efficiency about one percentage point below those used in §2.1 through §2.5. Both STIG and ISTIG
exhibit similar efficiencies and both are well below that of a combined cycle of similar GT technology.
During the 1980’s, STIG cycles were “in vogue” and some of their exponents published articles in peerreviewed academic journals stating that STIG cycles are fundamentally more efficient than combined
cycles!
220
52
Efficiency @ Gen. Term., %
51
330
440
550
660
770
880
990
Solid Lines: Non-intercooled
Dashed Lines: Intercooled
kJ/kg
/gt-00026/
50
PR = 12
PR = 16
PR = 24
49
48
PR = 32
47
46
45
TIT=1900-2700 F (1038-1482 C)
by increments of 200 F (111 C)
44
100
150
200
250
300
350
400
450
1700 F =
1900 F =
2100 F =
2300 F =
2500 F =
2700 F =
927 C
1038 C
1149 C
1260 C
1371 C
1482 C
Specific Power @ Gen. Term., kW per lb/s
Fig. 35. Performance of steam-injected cycles as a function of TIT and pressure ratio
2.8.2
DUAL-PRESSURE STEAM-INJECTION
To improve heat recovery with high pressure ratio engines, a second evaporator at a lower pressure is
added, downstream of the first in the exhaust gas path of the heat recovery boiler. This is analogous to a
dual-pressure combined cycle, discussed later in this work. The additional low pressure steam generated
cannot be introduced into the combustor, which is at a higher pressure, so it is introduced into the LP
section of the gas turbine. This cycle has been applied in the aeroderivative GE LM 5000 gas turbine,
known as a "STIG 120". Naturally, the secondary steam is less effective than the primary steam, since it
expands through a smaller turbine pressure ratio.
For the LM5000 engine, primary steam injection at 80,000 lb/hr (8.4% of nominal dry airflow) increases
power by 39% and efficiency by 15%. Further secondary injection of 40,000 lb/hr (4.2% of nominal dry
airflow) adds only 11% to the power and 3.6% to the efficiency. The secondary steam injection is
approximately half as effective as the primary. The relatively cool secondary steam injection has the
drawback of quenching the hotter gases in the turbine engine. For the LM5000 STIG120, mixing the
secondary steam cools the turbine gases entering the LP turbine by about 30 ºC (54 ºF).
The secondary steam injection provides very low benefits to the cycle, unless its use enables an increase
of T3, to compensate for the gas quenching effect, which is the case for the LM5000 engine. This is
because when primary steam is injected into the combustor of the LM5000, it tends to increase the RPM
2-34
 M. A. Elmasri, 1990-2007
Chapter 2: GT Cycle Thermodynamics
of the HP spool, which needs to be countered by a reduction in T3. However, if secondary steam is
injected at the inlet of the LP turbine, it increases the pressure behind the HP spool, which slows it down,
permitting greater primary injection without the need to reduce T3.
Thus, the benefit of the dual-pressure STIG cycle in the LM5000 application is not intrinsic to the
thermodynamic process, but a consequence of applying the cycle to an existing engine design, whilst
trying to keep its modifications to a minimum. If an engine were designed from scratch to provide an
optimum STIG cycle, it would not use a dual-pressure system.
Fig. 36 summarises the STIG cycle results, for the practical engines for which it has been applied. As a
cycle for pure power generation, it provides a combination of cost and efficiency that is of dubious merit.
Efficiency is well below a combined cycle and capital cost is not low enough., especially when one
considers water treatment costs. However, for cogeneration plants with variable process steam demand,
where one can produce and profitably sell extra power at times when less process steam is being used, the
STIG cycle can be economically attractive.
Allison 501 KH (Cheng)
No injection:
5 MW, η=30.5%
Max. Injection:
6.5 MW, η=38 %
GE LM5000
No Injection:
33 MW, η=36%
Max. Injection, one pressure (all at combustor, STIG 80):
46 MW, η=41.5%
Max. Injection, two pressures (combustor & LP T, STIG 120): 51 MW, η=43%
No injection, 3-P Unfired Combined Cycle:
45 MW, η=49.5%
•
High specific power
•
Efficiency much lower than a combined cycle
•
Cost/efficiency combination of dubious benefit for pure power generation
•
Good flexibility and economics in certain, unique cogen applications
Fig. 36. Summary of characteristics of commercially proven “STIG” or “Cheng” cycles. Note that the efficiencies
shown are “system efficiencies” and not true gas turbine efficiencies because the GT is charged for the fuel but not
for the energy of the steam it receives.
2-35
 M. A. Elmasri, 1990-2007
Chapter 2: GT Cycle Thermodynamics
References
[1] M. A. Elmasri, "GASCAN - An Interactive Code for Thermal Analysis of Gas Turbine Systems", J. of Eng. for Gas Turbines & Power, Vol. 110, pp. 201-209,
1988.
[2] M. A. Elmasri, "A Flexible, Efficient Gas Turbine Cogeneration Cycle with a Novel Dual-Mode Heat Recovery System", Proc. 2d. ASME Cogen-Turbo,
Montreaux, Switzerland, Sept. 1988, pp. 229-237.
[3] M. A. Elmasri, "Exergy Analysis and Optimization of Recuperated Gas Turbine Cycles", Symposium on Computer Aided Engineering of Energy Systems,
1986 ASME WAM, Anaheim, CA, AES-Vol. 2-1, pp. 125-134.
[4] "ICR Marine Propulsion System Makes Cruise Engines Obsolete", Gas Turbine World, July-August, 1985.
[5]. Mills, R. G. and K. W. Karstensen, "Intercooled/Recuperated Shipboard Generator Drive Engine", ASME 86-GT-203.
[6] Staudt, J.E. et. al., "Intercooled and Recuperated Dresser-Rand DC990 Gas Turbine Engine", ASME 89-GT-3.
[7] Lysholm, A., US patent 2,115,338; April, 1938.
[8] Gasparovic, N. and J. G. Hellemans, "Gas Turbines with Heat Exchanger and Water Injection in the Compressed Air", Proc. Instn. Mech. Engrs., 185, 66/71,
1971.
[9] Foote, W.R., US patent 2,869,324; January, 1959.
[10] Martinka, M., US patent 2,186,706; January, 1940; and German patent number 717,711; February, 1942.
[11] M. A. Elmasri, "A Modified, High-Efficiency, Recuperated Gas-Turbine Cycle", J. Eng. Gas Turbines & Power, Vol. 110, pp. 233-243, April 1988.
[12] M. A. Elmasri, US patent 4,753,068. June 1988.
[13] Mori, Y. et al, "A Highly Efficient Regenerative Gas Turbine System by New Method of Heat Recovery with Water Injection", 83-Tokyo-IGTC-38.
[14] Cheng, D.Y., US patent 4,128,994; Dec. 1978.
[15] Cheng, D.Y., US patent 4,297,841; Nov. 1981.
[16] Brown, D. H. and A. Cohn, "An Evaluation of Steam Injected Combustion Turbine Systems",Journal of Engineering for Power, 103, pp 13-19, 1981.
[17] Leibowitz, H. and E. Tabb, "The Integrated Approach to a Gas Turbine Topping Cycle Cogeneration System",J. Eng. for Gas Turbines & Power, 106, pp
731-736, 1984.
[18] Burnham, J. B., M.H. Giuliani and D.J. Moeller, "Development, Installation and Operating Results of a Steam Injection System (STIG) in a General Electric
LM5000 Gas Generator", ASME 86-GT-231.
[19] Digumarthi, R. and C.N. Chang, "Cheng-Cycle Implementation on a Small Gas Turbine Engine",J. Eng. for Gas Turbines & Power, 106, pp 609-702, 1984.
[20] Jones, J. L, B.R. Flynn and J.R. Strother, "Operating Flexibility and Economic Benefits of a Dual Fluid Cycle 501KB Gas Turbine Engine in Cogeneration
Applications", ASME 82-GT-298.
[21] Larson, E.D. and J.H.Williams, "Steam-Injected Gas-Turbines", J. Eng. for Gas Turbines & Power, Vol.109, pp. 55-63, 1987.
[22] Kolp, D. A. & D. J. Moeller, "World's First Full STIG LM5000 Installed at Simpson Paper Company", ASME 88-GT-198.
[23] Thames, J.M. & R.P. Coleman, "Preliminary Performance Estimates for a GE Steam Injected LM1600 Gas Turbine", ASME 89-GT-97.
[24] Ecob, D.J. & D.G. Marriott, "Development of Steam Injection for Ruston TB5000 and Tornado Engines", ASME 89-GT-253.
[25] Kolp, D.A., S.R. Gagnon & M.J. Rosenbluth, "Water Treatment and Mosture Separation in Steam-Injected Gas Turbines", ASME 90-GT-372.
[26] Kan, S., T. Morishita & K. Hiraoka, "Reheat Gas Turbine with Hydrogen Combustion between Blade Rows", 83-Tokyo IGTC-27.
[27] El-Masri, M.A., "On Thermodynamics of Gas Turbine Cycles - Part I: Second Law Analysis of Combined Cycles", J. of Eng. for Gas Turbines & Power, Vol.
107, No. 4, pp. 880-889, 1985.
[28] El-Masri, M.A. & J.H. Magnusson, "Thermodynamics of an Isothermal Gas Turbine Combined Cycle", J. of Eng. for Gas Turbines & Power, Vol. 106, No. 4,
pp. 743-749, 1984.
[29] Hori, A. and K. Takeya, "Outline of Plan for Advanced Research Gas Turbine", ASME 81-GT-28, Journal of Engineering for Power, 103, 4, 1981.
[30] Teshima, K. et. al. "A Report for the Engineering Status of Some Confirmation Tests in the Development of the AGTJ-100A", ASME 84-GT-53.
[31] Takeya, K., Y. Oteki and H. Yasui, "Current Status of Advanced Reheat Gas Turbine AGTJ-100A (Part 3) Experimental Results of Shop Tests", ASME 84GT-57.
[32] Arai, M. et. al., "Results from the High Temperature Turbine Tests on the HPT of the AGTJ 100-A", ASME 84-GT-235.
[33] El-Masri, M.A., "Exergy-Balance Analysis of the Reheat Gas Turbine Combined Cycle", Proc. 1987 ASME-JSME Thermal Engineering Joint Conference,
Honolulu, Hawaii, Vol. 2, pp. 117-126.
[34] El-Masri, M.A., "Thermodynamics and Performance Projections for Intercooled/Reheat/Recuperated Gas Turbine Systems", ASME 87-GT-108.
[35] Foster-Pegg, R.W., "Turbo-STIG - The Turbocharged Steam Injected Gas Turbine Cycle", ASME 89-GT-100.
2-36
 M. A. Elmasri, 1990-2000
Chapter 3: GT Components - Compressors
GAS TURBINE COMPONENTS - COMPRESSORS
Content Revised November, 2000
© Maher Elmasri 1990-2000
Figure 1a. Typical medium (85 MW) heavy-duty GT rotor & stator with axial compressor (GT11N
gas turbine, courtesy of Alstom Power)
Two common types of compressors are used in GT's: Axial and Centrifugal (or Radial). The latter are
best suited for relatively small flow rates, below about 15 lb/s, and are therefore common in machines of
3-1
 M. A. Elmasri, 1990-2000
Chapter 3: GT Components - Compressors
below 1.5 MW. Many of the principals are the same, but we shall concentrate on the axial type, which is
predominant in larger machines. Figures 1a and 1b show typical heavy-duty gas turbines with axial
compressors.
Figure 1b. Typical large (250 MW) heavy-duty GT rotor & stator with axial compressor (501-G
gas turbine, courtesy of Siemens-Westinghouse)
1. FUNDAMENTALS OF A COMPRESSOR STAGE
1.a. Application of Conservation Laws of Momentum & Energy:
Fig. 2 illustrates a compressor stage. The absolute velocity of air entering the rotor is c1 with a tangential
component of cy1 in the rotor's direction. The flow is turned in the rotor, so the exiting air has absolute
velocity c2 with a tangential component cy2. The rate of change of angular momentum of air as it passes
through the rotor must equal the torque τ,
τ = m r (cy2 - cy1) ..................…………..... (1)
where r is the radius and m the air mass flow rate. The power P applied to the rotor is the product of
torque and angular velocity, in radians/s, ω so that
3-2
 M. A. Elmasri, 1990-2000
Chapter 3: GT Components - Compressors
P = m ω r (cy2 -cy1) .....…………………......... (2)
and the work done per unit mass of air, which must equal the increase in air stagnation enthalpy ho, is
W = u (cy2 - cy1) = (h02 - h01)= (h03 - h01) ........... (3)
where u is the tangential speed of the rotor, u = ω r. The subscript o implies the stagnation or total
thermodynamic property. This is the value the property would attain if the flow were brought to rest,
recovering all its kinetic energy. This "bringing to rest" would have to be isentropic to attain the
stagnation pressure, but does not have to be so to attain the stagnation temperature and enthalpy. The last
equality in eq. (3) is because the stator does no work, so conservation of energy implies a constant
stagnation enthalpy across it (i.e. h03 = h02). Eq. 3 is a simple statement of the laws of conservation of
momentum and energy. It is valid whether the flow is compressible or incompressible, and whether the
flow is isentropic (reversible) or not.
β1
c1
w1
w1
ROTOR
u=ωr
w2
cx1
u
c1
cy1
α2
β2
c2
α1
c2
cx2
w2
u
STATOR
cy2
c = absolute velocities
c3
w = relative velocities
Figure 2. Basic aerodynamics of a compressor stage
1.b. Pressure Rise in a Stage:
The pressure rise corresponding to the enthalpy rise of eq. 3 depends on stage efficiency. If the stage
were perfect, with totally frictionless flow, that pressure rise would be the maximum; all the work input
would go into increasing the exergy of the air, and would be fully recoverable by perfectly re-expanding
the air. If the stage is totally inefficient, the flow will be "churned-up" and just heated at constant
stagnation pressure, all the work input will be converted to thermal energy and much of the exergy
destroyed. Only part of the work of compression would then be recoverable from the heated, lowpressure air, with the portion of that heat recoverable as work increasing with temperature difference from
the ambient.
3-3
 M. A. Elmasri, 1990-2000
Chapter 3: GT Components - Compressors
Ideal Stage:
For the perfectly efficient limit, if we assume a perfect gas, the stagnation (or total) enthalpy is directly
proportional to stagnation (total) temperature, and the isentropic pressure-temperature relationship can be
used to find the pressure ratio of the stage as:
( h03,s / h01 ) = ( T03,s / T01 ) = ( P03 / P01 )(k-1)/k
...... (4)
The subscript s means "isentropic" in the above equation.
Real Stage:
The stage adiabatic efficiency, η, also called isentropic efficiency, is defined as the ratio (Isentropic
Enthalpy Rise / Actual Enthalpy Rise) corresponding to a certain increase in pressure. This may be
defined as "total-to-total", if based on the inlet and exit stagnation states; "total-to-static", if based on the
inlet stagnation state and exit static state; or "static-to-static", if based on the inlet and exit static states. In
the present discussions, the "total-to-total" is efficiency definition is used, as illustrated on Fig. 3.
h03,s - h01
η = -------------h03 - h01
................... (5)
Figure 3. Isentropic and real adiabatic compression
h or T
p03
03
03s
Real
Compression
Isentropic
Compression
p01
01
s
The state-of-the-art in compressor blading design defines how efficient a single stage can be. A typical
stage has a small pressure ratio, about 1.2. Thus, it is convenient to define efficiency of a compression
process on the basis of that for a small stage, called the Polytropic Efficiency , Є
Є = dh0,s/dh0 .......................... (6)
3-4
 M. A. Elmasri, 1990-2000
Chapter 3: GT Components - Compressors
where the small increases in enthalpy, temperature and pressure across the stage relate the exit to the inlet
values by
h03,s = h01 + dh0,s
h03 = h01 + dh0
T03,s = T01 + dT0,s
T03 = T01 + dT0
p03 = p01 + dp0
.................... (7)
so that by using the binomial theorem, eq. 4 may be reduced to
(dh0,s / h0) = (dT0,s / T0) = [ (k-1)/k ] (dp0 / p0)
……… (8)
and eq. 6 gives
(dh0 / h0) = (dT0 / T0) = [ (k-1)/k Є ] (dp0 / p0) ……… (9)
where the perfect gas assumption was used to linearly relate enthalpy and temperature. Fig. 4 shows the
T-s diagram for compression of a perfect gas.
Figure 4 Polytropic, or "small-stage" compression
h or T
p0 + dp0
h0+dh0
h0+dh0,s
p0
h0
s
Eq. 9 can be integrated for a finite, real stage, or group of stages with a constant polytropic efficiency, to
give
( h03 / h01 ) = ( T03 / T01 ) = ( P03 / P01 )(k-1)/kЄ
........ (10)
Compounding a series of small stages, results in an overall efficiency smaller than the polytropic, because
any inefficiency in the upstream stages results in hotter air reaching the downstream stages. Compressing
this hotter air requires more work per unit mass than compressing the cooler air which would have
3-5
 M. A. Elmasri, 1990-2000
Chapter 3: GT Components - Compressors
reached them had the preceding stages been more efficient. Thus, for a given polytropic efficiency,
representing a given state-of-the-art, the adiabatic efficiency falls with increasing pressure ratio, as seen in
Fig. 5.
Interestingly, if one defines a second law efficiency as the ratio of (increased exergy of air)/(work input),
it can be shown to increase with pressure ratio. The reason is that work dissipated into heat near the
beginning of compression is at low temperature, less convertible to work than high temperature heat
resulting from dissipation at the higher temperature near the end of the compression. Fig. 5 shows this
result as well.
Overall and Exergy Efficiencies of a Compression
with a Polytropic (small stage) Efficiency of 0.9
0.96
Efficiencies
0.94
0.92
0.9
Overall efficiency
0.88
Exergy efficiency
0.86
0.84
0.82
0
10
20
30
40
Pressure Ratio
Figure 5
1.c. Relation between Stage Work and Blade Velocity & Angles:
For the case cx1=cx2, a constant axial velocity, eq. 3 can be combined with trigonometric relations from
Fig. 2 to give
(h03 - h01) = u2 { 1 - (tan β2 +tan α1 ) cx/u } …….. (11)
i.e. the stage work is proportional to the square of the peripheral speed and to the extent the flow is turned
by the blades.
The ratio (h03 - h01)/u2 is defined as the stage loading coefficient, ψ, and the ratio cx/u as the flow
coefficient, φ, so that in dimensionless form, eq. 11 is written as
ψ = 1 - φ (tan β2 + tan α1)
………….…..… (11a)
Referring to the velocity triangles of Fig. 2, an observer moving with the rotor sees air coming in at the
relative velocity w1 and leaving at the relative velocity w2. In that relative coordinate frame, the rotor
appears stationary, doing no work. Thus the rotor-relative stagnation enthalpy, the sum of static enthalpy
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 M. A. Elmasri, 1990-2000
Chapter 3: GT Components - Compressors
and relative kinetic energy, is constant,
h1 + w12 /2 = h2 + w22 /2
...……….......... (12)
so the static enthalpy increases across the rotor if there is a deceleration of relative velocity,
h2 - h1 = ( w12 - w22) /2 .……………......... (13)
thus in the relative frame, the rotor behaves as a diffuser, reducing relative kinetic energy and building up
static enthalpy, as well as temperature and pressure. Apart from this pressure buildup, the rotor also
increases the absolute air velocity, so the leaving velocity c2 is greater than the entering velocity c1. This
permits the next stator to further increase static enthalpy, temperature, and pressure at the expense of
reducing kinetic energy,
h3 - h2 = ( c22 - c32 ) /2 ..……………........ (14)
Normally, a stage is designed with c3 = c1 so that the next stage has identical triangles. It is also common
to design with the axial component of velocity, cx, essentially constant through the stage and varying
slowly from stage to stage.
Low Pressure
High
Velocity
Mainstream
Slow, Viscous
Boundary
Layer
High Pressure
Figure 6
Both sets of blades act as diffusers in their respective coordinate frames, designed to decelerate the air,
converting its kinetic energy to pressure. Thus the static pressure, temperature and enthalpy increase at
the expense of falling velocity. The paradox that the velocity is continuously "falling", relative to each
blade passage, yet is the same after the stage, is a consequence of the relative motion of the blade
passages. The velocity in the absolute frame rises through the rotor and falls through the stator. In the
blade-relative frame it falls throughout.
In the blade-relative frame, no work is done on the flow, and stagnation temperature and enthalpy, based
3-7
 M. A. Elmasri, 1990-2000
Chapter 3: GT Components - Compressors
on relative velocities, are constant. Relative stagnation pressure is constant across a frictionless,
isentropic blade-row but falls a few percent in a real case. In the absolute, stationary frame, the rotor does
work and creates the entire rise in the stagnation state.
Since both sets of blades act as diffusers, they operate with relative velocity declining and static pressure
increasing along the airflow direction. The air flowing near the blade surfaces forms a "boundary layer"
retarded by viscous friction. This could leave it with insufficient momentum to carry it into the adverse
pressure gradient. If this boundary layer separates, Fig. 6, it could lead to a blade stall, with a sudden loss
of pressure-raising capability. If that progresses throughout a multi-stage compressor, the entire
compressor would "surge" and compressed air (and fire in a GT engine) at its delivery-end could blow
back through its front end. The compressor must therefore be designed with a moderate rate of increase
of pressure per blade row.
To maximize the pressure rise in a stage, it should be approximately equal for rotor and stator. A reaction
parameter , or degree of reaction, is defined as the ratio of static enthalpy rise in the rotor to the stage
total enthalpy rise:
R = (h2 - h1) / (h03 - h01) ...........…………........... (15)
and if c3 = c1, then
R = (h2 - h1) / (h3 - h1).
A degree of reaction of 50% implies equal apportionment of static enthalpy rise between rotor and stator.
It also implies identical velocity triangles if cx is constant. This simplifies compressor design
considerably since all blade rows would be mirror images of each other, decreasing in annulus height and
blade chord as the air gets compressed into a smaller volume. A reaction parameter of about 50% is
commonly used.
Zero reaction corresponds to a situation where the rotor just accelerates the flow, at constant static
pressure, and the stator does all the diffusion, converting kinetic energy to enthalpy and pressure increase.
This implies a constant rotor-passage flow cross section. Such a design would have the advantage of no
leakage across the rotor blade tips since there would be no pressure difference across the rotor. However,
an excessive number of stages would be needed since the limited rate of diffusion is concentrated in
stators only. 100 % reaction would be the opposite extreme, with all pressure rise in the rotors, the stators
being used only to redirect the flow at constant velocity and pressure. Such a design would have no
merit.
2. VARIABLE INLET GUIDE VANES (IGV'S)
To reduce airflow, a variable inlet guide vane can be installed at the inlet to the first stage. Fig. 7
illustrates its principles, and Fig. 8 shows a typical construction. Closing the IGV imparts a swirl to the
air in the rotor's direction. The volumetric flow rate is reduced, as can be seen from the reduced axial
component, cx. The increase in swirl component, from cy1 to cy1' reduces the first-stage work by reducing
the rate of change of angular momentum from (cy2 - cy1) as seen from eq. 3. Thus the pressure, density,
and mass flow rate are reduced at the inlet to the second and subsequent stages.
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 M. A. Elmasri, 1990-2000
Chapter 3: GT Components - Compressors
Inlet Guide Vanes
c1
β1
c1 '
α1
w1
c1
u
w1
ROTO
u=ωr
w2
β2
w2
c2
cy1
cy1'
α2
c2
cx2
u
Fig. 7
cy2
Fig. 8 Variable IGV ring (501F gas turbine, courtesy of Siemens-Westinghouse)
3-9
 M. A. Elmasri, 1990-2000
Chapter 3: GT Components - Compressors
3. TRANSONIC COMPRESSORS
For a given geometry, eq. 11 shows that stage work is proportional to the square of velocities. Using a
higher tangential speed can therefore reduce the number of stages. If the tangential speed is increased
beyond a certain point, the relative velocity entering the rotor (w1 in Fig. 2) can become supersonic. This
happens at the first stage where the air temperature is coolest and the speed of sound lowest. If the blade
passages are designed for subsonic flow, a shock will stand in front of the rotor and a distorted threedimensional flow pattern will develop. Efficiency would fall sharply and the stage may stall. Recently,
compressors are being designed to operate with transonic first stage blades, where the velocity would be
supersonic at the tips but subsonic at the hubs. Highly complex blade shapes are needed for that purpose.
They operate satisfactorily, at a lower efficiency than subsonic stages, but the resulting high pressure ratio
reduces the number of stages, compressor length and cost.
With the largest gas turbines currently used for power generation, transonic first stages are imposed by
the rotor diameter and the constraint that it is unrealistic to operate at an RPM below synchronous speed.
4. FLOW PATH ANNULUS TAPER AND STARTING
Since air density increases towards the rear of a compressor spool, maintaining a reasonable axial velocity
requires a progressively decreasing annulus area. For a typical heavy duty GT, the air at compressor
discharge is about seven times as dense as the air at the inlet, so maintaining a uniform air axial velocity,
with a constant mean diameter, would require the blades of the last row to be only 15% of the length of
those of the first blade row.
Reducing the annulus area may be accomplished by various methods. Maintaining the tip diameter
constant leads to increasing hub and pitch diameters and increasing pitch velocity (u). It has the
advantage of increasing the stage work and reducing the number of stages needed. Sonic velocity
increases with temperature and so the flow Mach Number will decrease towards the rear despite the
increasing pitch speed. Constant tip diameter has the disadvantage that shorter blades are needed for a
given annulus area, increasing tip-clearance in proportion to flow-path area, which increases leakage; and
increasing end-wall area in proportion to flow-path area, which increases end-wall friction. Maintaining
the hub diameter constant reverses the advantages and disadvantages above. Maintaining pitch diameter
constant provides a reasonable compromise. In modern machines, it is not uncommon for all three
diameters to be varied.
The limits on the maximum hub/tip diameter ratio at the discharge, and on minimum hub/tip ratio at the
inlet (to avoid excessive blade twist and three dimensional losses) result in a maximum, practical pressure
ratio on a single spool. By historic standards, this maximum pressure ratio was considered to be about
18:1, but recent designs, such as the Alstom GT24/26 have a single spool compressor with a ratio of
about 30:1.
Aeroderivative gas turbines have historically had high pressure ratios, and have employed multi-spool
compressors to reduce flow area as the air is compressed. The rear portion of the compressor would be
designed with a smaller diameter to reduce annulus area without resorting to very short blades in
proportion to pitch diameter; and would rotate at a higher speed to produce reasonable blade speed (u),
flow velocity (cx), and flow coefficient (φ = cx /u).
3-10
 M. A. Elmasri, 1990-2000
Chapter 3: GT Components - Compressors
Figure 9. Intermediate bleeds in a GT compressor (GT11N, courtesy of Alstom Power)
When operating at the design condition, all the blades on a spool will have proper velocity triangles. The
annulus height taper will keep the axial velocity component at or near its design value for each stage by
reducing the flow area towards the rear of the compressor as air density increases.
When the compressor starts, it is filled with air at atmospheric density, which begins to flow through the
machine. The small flow areas at the back of the compressor will block the flow by choking and the axial
velocity at the front stages will therefore be too low, stalling the front stages. This prevents the pressure
from building up and the proper density distribution cannot be established. To relieve the front stages
from blockage by the rear ones, one or more intermediate bleed ports are opened at startup, as shown in
Fig. 9. This enables adequate flow velocities at the front stages, buildup of pressure, reduction of density
at the rear stages, and commencement of operation. The bleed ports are closed as the velocity, pressure,
and density distribution become established.
5. THE COMPRESSOR MAP
The compressor is a pumping device, analogous to a centrifugal pump, with which most mechanical
engineers are more familiar. A pumping device is tested by running it at a given speed, then varying the
resistance against which it works, such as by placing a valve or damper at its exit. For a given rotational
speed N, the resistance of the valve or damper determines the pressure and flow. Shutting the valve or
damper allows one to trace the pressure-flow characteristic, as shown by the points a-b-c in Fig. 10 below.
Running the pump or compressor at a different speed then allows one to trace the pressure-flow
3-11
 M. A. Elmasri, 1990-2000
Chapter 3: GT Components - Compressors
characteristic at another speed, such as N', (lower than N in the sketch of Fig. 10).
N
Delivery
pressure
Axial
compressor
Delivery
pressure
c
Surge line
Centrifugal
pump
c
b
b
closing
damper
N'<N
a
N'<N
closing
valve
a
N
N
Flow rate
Flow rate
Figure 10. Pressure-flow characteristics (pumping curves) of typical turbomachines
Compounding many stages in series in an axial compressor results in a very steep pressure-flow curve.
This is because a small decrease in volumetric flow at the first stage causes it to produce a higher exit
pressure, which, due to the increased density, causes an even greater reduction in volumetric flow to the
second stage, which therefore produces an even greater pressure rise. Thus compounding the stages
amplifies the effect of a very small change in inlet volumetric flow rate, producing a steep curve, as
shown in Fig. 10.
If the valve is completely shut on a centrifugal pump, point c in Fig. 10, it still produces a pressure rise
corresponding to the centrifugal force acting on the spinning ring of trapped water. An axial compressor,
however, produces pressure from the flow velocity through the blade passages, which act as diffusers.
Thus, if the flow is blocked, no pressure will be produced, and the blades will just churn up the trapped
air and heat it without building any pressure. As a matter of fact, as the damper behind the compressor is
closed along the curve a-b-c, the ability to compress will disappear long before the flow rate reaches zero.
Once the flow velocities through the blade passages are slowed down to the point where the mainstream
is unable to help propel the viscous boundary layer into the adverse pressure gradient, the flow will stall,
with the boundary layer separating and destroying the smooth streamlined flow through the diffuser
3-12
 M. A. Elmasri, 1990-2000
Chapter 3: GT Components - Compressors
passages, as depicted in Fig. 6. When this point, c in Fig. 10, is reached, the compressor will suddenly
and catastrophically lose its ability to compress. This condition is known as compressor surge. As it is
approached during a test, the pressure at the compressor exit will begin to fluctuate, and the machine will
begin to vibrate, since small local blade stalling will begin to occur on and off, with the machine still
compressing. If one continues to close the damper, these fluctuations become more violent in amplitude,
and the maximum pressure ratio which can be achieved at a given speed is marked by point c. The
compressor's normal operation is then set at a sufficient margin below this point. This point is found for
each speed, and a surge line is marked on the compressor map.
Fig. 11 illustrates a typical compressor map. Experimentally, the map is generated by running the
compressor at different RPM's and modulating a downstream valve to trace the pressure-massflow curve.
For each RPM in the normal operating range, the massflow is virtually independent of pressure ratio, and
the constant-RPM lines are nearly vertical. As the compressor is back-pressured towards its limit, the
mass flow begins to fall more pronouncedly, and shortly thereafter the compressor will surge. Surge is a
complex process that is not perfectly predictable by analytical or numerical methods. It starts by stalling
one or more blades in a row, and the stall cell may rotate circumferentially around the affected stage,
periodically stalling and unstalling each blade passage, whilst other stages function more or less normally.
As the condition gets aggravated, the stall will propagate to other blade rows. Some compressors would
operate with partially-stalled blade rows in an unstable mode and recover as the back pressure is reduced.
Others could suffer damage from flow-induced vibrations of the blading. In all cases, it is advisable to
stay well away from the surge line by a suitable margin.
To generalize the results, the data is non-dimensionalized. From equations 4, 5 and 11, one may write the
pressure ratio of the first stage as
PR = { 1 + η Mu2 (k-1) [1 - (tan β2 + tan α1) Mx/Mu ] }k/(k-1)
which shows that, for a given geometry and air specific heat ratio, the pressure ratio is primarily a
function of the stage's axial Mach Number and blade tangential speed Mach Number. Both of these Mach
Numbers at any stage are essentially proportional to their equivalents at the first stage. If variable
geometry is used, such as variable IGV's, several compressor maps must be drawn, one map for each
geometric setting, at suitable intervals, to allow interpolation between maps.
The stage pressure ratio also depends on its efficiency. Stage efficiency is a dependent variable, function
of the flow field and pressure ratio. The inviscid flow field depends only on the geometry and Mach
Numbers, but the viscous effects depend on the Reynolds Number as well. The stage pressure ratio is
therefore a function of geometry and three independent dimensionless variables: Reynolds Number and
the axial and tangential Mach Numbers. For a given compressor, the Reynolds Number at each point
within the compressor scales primarily with the Mach Numbers, with some additional dependence on the
ratio of inlet pressure to inlet temperature, which varies by +/- 15% across a typical operating range of a
machine. Since the Reynolds Number effects tend to scale with Re0.2, the effect of Re as an independent
variable is quite small for a given machine (although it can become significant in scaling the map from
one machine to another of different size, since a smaller machine has a lower Reynolds Number, implying
greater viscous effects and lower efficiencies).
Thus, the pressure ratio of a particular compressor can be mapped out as a function of the primary
independent variables which control its performance, the axial and tangential Mach Numbers (for each
geometry if variable). To facilitate using the resulting map, those Mach Numbers are recast as
dimensional variables that are directly proportional to the dimensionless ones. From the continuity
equation (m = ρ Aan cx) and the equation of state ( p = ρ R T ),
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 M. A. Elmasri, 1990-2000
Chapter 3: GT Components - Compressors
cx = (m / ρ Aan) = ( m R T / p Aan)
thus, the axial Mach Numbers at each stage are directly proportional to the Flow Function at the first
stage (m Ta1/2 / pa), which has complex dimensions, and is therefore conveniently recast as a
Corrected Flow, that has physical dimensions of lb/s or kg/s by dividing the inlet air temperature by a
constant reference value Tref, usually 518.7 ºR (59 ºF) [288.15 K (15 ºC)]; and dividing the inlet air
pressure by a constant reference value, pref, usually 14.696 psia [1.013 bar], thus
Mx = cx/(kRTa)1/2 ~ m Ta1/2 / pa ~ m (Ta/Tref)1/2 (pref/pa)
where m (Ta/Tref)1/2 (pref/pa) is the Corrected Flow, customarily used as the x-axis on the
compressor map.
14
Margin
Pressure
Ratio
Surge Line
12
88%
10
85%
8
6
η = 83%
5100
5900
Corrected speed,
4700
N
Ta
Tref
4300
3900
220
5500
6300
Choke Line
240
260
280
300
Corrected flow, (lb/s)
m
Ta
pa
Tref
320
340
pref
Fig. 11. Typical compressor map
Likewise, the tangential Mach Number, Mu, is directly proportional to the Corrected Speed
Mu = u/(kRTa)1/2 ~ N (Tref/Ta)1/2
where N is the rotational speed (RPM) of the compressor. Corrected speed, N (Tref/Ta)1/2 , is therefore
used as the speed parameter on the compressor map.
3-14
 M. A. Elmasri, 1990-2000
Chapter 3: GT Components - Compressors
6. MATCHING THE COMPRESSOR AND TURBINE
Whenever a pumping device feeds a flow-resisting device, the operating point of the system is set by the
intersection between the pumping curve and the resistance curve. Fig. 12 illustrates this principle, both
for a common hydraulic system of a pump feeding a pipeline, and for the present system of a compressor
feeding a turbine through a combustor. The turbine behaves substantially like a choked nozzle, so that the
pressure required at its inlet, to force a certain gas flow through, is essentially proportional to the mole
flow rate and the square root of the temperature :
where
pturbine inlet = C Anozzle nt Tinlet1/2 f(MW, k)
C
Anozzle
nt
Tinlet
f(MW, k)
is a constant
is the turbine inlet nozzle cross-sectional area at the throat
is the mole flow rate of gases entering the turbine (mass flow/molecular weight)
is the turbine inlet temperature
is a secondary correction, function of molecular weight and specific heat ratio
Axial
compressor
Centrifugal
pump
typical parabolic
resistance curve
Delivery
p ~ m2
pressure
typical turbine
resistance curve
p ~ nt Tinlet1/2
Delivery
pressure
operating
point
operating
point
N'<N
N
N
N'<N
Compressor Inlet Mass Flow rate
Mass Flow rate
Figure 12. Balance between pumping and resistance curves determines operating point
Thus, a turbine with a fixed nozzle area, running at the nearly-constant full-load firing temperature,
without steam or water injection, will offer a flow resistance in proportion to the compressor inlet flow
rate, resulting in the line shown in Fig. 12. At such full-load operation, the compressor inlet flow rate will
vary with ambient temperature. Since air is denser on a cold day, the compressor will ingest a greater
mass flow, so the turbine will demand a proportionate increase in its inlet pressure.
A typical heavy-duty, single-shaft gas turbine runs at constant speed. Thus, its corrected speed increases
on a cold day, and falls on a hot day, in inverse proportion to the square root of absolute inlet temperature.
3-15
 M. A. Elmasri, 1990-2000
Chapter 3: GT Components - Compressors
For example a machine designed to run at 5100 RPM will also have a corrected speed of 5100 RPM at
ISO conditions (518.67 ºR or 288.15 K). On a cold day, at 0 ºF (-18º C) say, its real, physical speed is the
same, so its corrected speed is 5100 (518.67/459.67)1/2 = 5417 RPM. On a hot day, at 100 ºF (38º C) say,
its real, physical speed is the same, so its corrected speed is 5100 (518.67/559.67)1/2 = 4910 RPM.
Thus, across the typical range of ambients commonly encountered, a heavy-duty gas turbine's corrected
speed will vary by ± 5%, its physical airflow will vary by ± 10%, and its corrected airflow will vary by ±
5%. The variation of physical flow will result in an equivalent variation of pressure ratio at full-load of
roughly ± 10%. Running at part-load involves reducing airflow via the IGV's, or reducing firing
temperature, or both. Thus pressure ratio will fall at part-load. Injecting steam or water into the
combustor, or using a fuel with a low calorific value (and correspondingly a larger mass flow rate than
normal fuel) will increase pressure ratio since a larger mole flow of hot gases need to be forced through
the turbine. Fig. 13 illustrates the operating range on a typical compressor map.
14
Margin
Full-load
operating
line (steam
injection 5%)
Pressure
Ratio
Cold day
Surge Line
12
88%
10
Full-load
operating
line (dry)
85%
Hot day
8
6
η = 83%
5100
4700
5900
Corrected speed,
N
Ta
4300
3900
220
5500
6300
Choke Line
240
260
280
Corrected flow, (lb/s)
300
m
Ta
pa
Tref
320
pref
Figure 13. Typical full-load operating range for a single-shaft gas turbine
3-16
340
Tref
 M. A. Elmasri, 1990-2002
Chapter 4: GT Components - Combustors
GAS TURBINE COMPONENTS – COMBUSTORS
Content Revised May, 2002
© Maher Elmasri 2000 - 2002
1. FUNDAMENTALS OF GT COMBUSTORS
1.a. Introduction:
A typical gas turbine, burning fuel with a heating value of about 19,000 BTU/lb (44,000 kJ/kg), will have
an overall air/fuel massflow ratio of about 50:1. This ratio is over three times stoichiometric, and would
not sustain combustion if the fuel were mixed into the entire air stream.
The air velocity leaving the last stage of an axial compressor is about 500 ft/s (150 m/s), too fast to
sustain stable combustion. Even if that were accomplished, the resulting loss in stagnation pressure
would be exorbitant, as shown later. A diffuser is therefore used to bring the mean air velocity down to
the neighborhood of 80 ft/s (25 m/s) before delivering the air to the combustion chamber.
Figure 1. Simple throughflow combustor
Before NOx became a concern, the primary function of a combustor was to burn the fuel in a safe, stable
and reliable fashion. This was accomplished as illustrated in Fig. 1. Swirling airflow enters the region
around the fuel nozzle in an amount that produces a stable flame, i.e. the air/fuel ratio is just above
stoichiometric, with about 20 % excess air. This region, known as the "Primary Zone", must therefore
have roughly one-third of the engine's airflow. Combustion is essentially completed in this zone,
resulting in a hot fireball (about 3500 ºF/1930 ºC). The fireball is stabilized by creating a vortex through
the introduction of secondary air jets just beyond the primary zone. The recirculation of combustion
products enhances complete combustion. The combustion products are then be diluted by the remaining
air, bringing them it down to the 1st stage nozzle inlet temperature. The primary zone design must also
allow adequate turbulent mixing of air and fuel, and the mean through flow velocity must be slow enough
to provide sufficient residence time for the reaction.
A combustor liner made of thin heat-resistant alloy sheet metal, as illustrated in Fig. 1, is commonly used
to fulfill the above requirements. It controls the airflow to the primary zone as well as the introduction of
4-1
 M. A. Elmasri, 1990-2002
Chapter 4: GT Components - Combustors
secondary dilution air. It also serves to contain the very hot primary-zone fireball, shielding the casing of
the machine from contact with the fire as well as radiation from it. The secondary air flows around the
liner to cool it before entering into the liner for dilution. Some secondary air is introduced through small
corrugated slots along the liner, to flow in a cooling film along its inner wall surface, further reducing the
liner temperature by shielding it from the hotter gas.
As gas turbine firing temperatures and pressure ratios are increased, the burden on liner design gets
heavier. A bigger percentage of the total engine airflow is needed in the primary zone to cope with the
increased fuel. Also, a larger proportion of the engine's airflow, as much as 20%, is used for turbine
cooling, by-passing the combustion section altogether. Thus, less secondary air is available for dilution
and cooling the liner and transition piece. The compressor discharge is hotter with higher pressure ratios,
further complicating the problem. The level of sophistication in designing liner cooling techniques is
therefore continuously progressing. Ceramic thermal barrier coatings are used in some designs on the
inner surfaces, to bring liner metal temperatures closer to those of the cooling air.
In some recent designs, the concept of a sheet metal combustion liner has been abandoned, in favour of an
array of ceramic tiles lining the combustion section casing, with cooling air jets introduced onto the backsurface of the tiles.
1.b. Pattern Factor & Profile Factor :
Despite the turbulent mixing in the combustion and dilution processes, the temperature reaching the
turbine is not uniform. Fig. 2 illustrates a typical temperature profile reaching the turbine.
The "Pattern Factor"
P.F. = ( Tmax – TIT ) / ( TIT - TCD )
describes the maximum temperature relative to the mixed-mean TIT. Ideally for the stator, this would be
zero, so that the stator blades will only see the TIT. Actually, this varies between 0.1 and 0.2, which
means that the stator "hot-spots" may be well above the temperature corresponding to a uniform mean
TIT. The pattern factor can be reduced by:
(i) A long combustor liner & transition piece.
(ii) More vigorous mixing of secondary air. This requires smaller dilution holes in the liner and a larger
pressure drop for the airflow.
(iii) A large number of small burners.
The "Profile Factor" is the same as the pattern factor, except that it is based on the maximum
circumferentially-averaged temperature. This is most relevant to the rotor, whose temperatures might be
more restrictive than the stator's due to its high stress. In modern engines, the ideal profile factor is not
necessarily zero. Rather, it is a curve with a peak at a radius above the mean, where the centrifugal stress
on the blade is lower than near the root. This takes full advantage the material's capability to maximize
mean gas temperature.
4-2
 M. A. Elmasri, 1990-2002
Chapter 4: GT Components - Combustors
r
rtip
Tmax
rhub
P.F. = (Tmax – TIT) / (TIT – TCD)
T
TCD
TIT
Figure 2
1.c. Combustor Efficiency & Pressure Drop
Most combustors are close to 100 % efficient when operating normally, and the fuel is all burnt.
However, it is customary when analyzing gas turbine cycles to assign a combustion efficiency, between
99% and 99.8%. The 0.2 % to 1% loss is used to account for any unburnt fuel, but is largely a way of
factoring in a heat loss from the combustor as well as the rest of the engine.
Stagnation pressure drop in the combustion section of a GT consists of three components:
(i) Loss in the diffuser between compressor and combustor. Typically the compressor discharge flow is at
about 500 ft/s axial velocity, corresponding to a Mach Number of about 0.3. The diffuser has to slow it to
about one-fifth of those numbers. Even with a well-designed diffuser, about 1% to 1.5% of the stagnation
pressure is lost.
(ii) Loss due to heat addition. Heating an air stream expands and accelerates it. Applying the laws of
momentum and energy to a heated constant-area duct shows that stagnation pressure is lost, typically just
under 1% in a gas turbine combustor. Since that loss is proportional to the square of the flow velocity, it
is imperative to reduce velocity before combustion.
(iii) Loss in flow turning and across the liner. The former depends on type of combustor, being highest
for reverse-flow cans. The latter is necessary to force jets of air into the combustion and dilution zones at
sufficient velocity to promote mixing. This loss is typically 1.5% to 2.5% of stagnation pressure.
The total pressure loss is typically in the range 2.5% to 5% of stagnation pressure. Silo combustors tend
to have the lowest pressure losses.
4-3
 M. A. Elmasri, 1990-2002
Chapter 4: GT Components - Combustors
2. TYPES OF COMBUSTORS
There are basically four types of combustors:
a. Through-flow can.
b. Reverse-flow can.
c. Silo.
d. Annular.
2.a. Through-flow can combustors:
Figure 3. Aeroderivative with through-flow can combustors (FT8, courtesy of Pratt & Whitney)
These were used on most early jet engines, and continue to be used in modern designs, such as shown in
Fig. 3. A series of cylindrical liners, with a cross-section as shown in Fig. 1, are arranged around the
machines circumfrence. A transition piece, such as shown in Fig. 7, is used to connect the exit from the
circular can to the first-stage nozzle sectors.
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 M. A. Elmasri, 1990-2002
Chapter 4: GT Components - Combustors
With a several through-flow cans, there is more liner surface area to cool than in an annular combustor of
the same flow cross-section. Transition pieces with complicated geometry are still needed. The main
advantages of multiple through-flow cans is the ability to replace one at a time, and the ability to evolve a
design from an older, proven one and to test it economically with one can.
2.b. Reverse-flow can combustors:
Figure 4. Heavy-duty GT's with reverse-flow can combustors (Top: Frame 6B, bottom: Frame 9E,
courtesy GE)
This is the most common design in large stationery gas turbines. It is similar to the throughflow can,
except for the following advantages:
(i) Ability to lengthen the combustor liner to increase residence time without excessively lengthening the
machine shaft between compressor and turbine.
(ii) Possibility to disassemble and replace fuel nozzles, liners and transition pieces without disassembling
the machine.
(iii) Improved cooling of liners and transition pieces. The entire compressor discharge flow goes around
the latter, and then in counterflow around the former. The minimum airflow around the liner, in the
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 M. A. Elmasri, 1990-2002
Chapter 4: GT Components - Combustors
region of the primary zone and fuel nozzles, is almost one-third of compressor delivery. By contrast, the
minimum airflow around the transition pieces for annular and through-flow cans is towards the turbine
nozzles, where it approaches zero.
Figure 5. Heavy-duty GT with reverse-flow can combustors (501G, courtesy SiemensWestinghouse)
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 M. A. Elmasri, 1990-2002
Chapter 4: GT Components - Combustors
Figure 6. Combustor cans (above) and holding sleeves (below) for a heavy-duty GT with reverseflow can combustors (Frame 7E, courtesy GE & MassPower)
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 M. A. Elmasri, 1990-2002
Chapter 4: GT Components - Combustors
Fig. 6 shows combustor cans for an 85 MW heavy duty GT. The large holes in the cans are for cross-fire
tubes to propagate flame from one can to the next during startup and to relight a can if it suffers a
momentary flameout. The corrugations contain myriad small holes to introduce cooling air along the
inner surface of the cans. The combustor cans need to be secured with some freedom for expansion, and
in a manner that minimizes transmitting their vibration to the surrounding components and structure.
Sleeves with stand-offs that support the can and center it within the sleeve, while allowing it some
freedom to expand, are used for that purpose.
Figure 7. Transition pieces for a heavy-duty GT with reverse-flow can combustors (Frame 7EA,
courtesy GE and MassPower)
The transition piece is a duct that conveys the combustion gases, after dilution, into the turbine's firststage nozzles. Its shape depends on the type of combustor. In annular combustors, typical of modern
aero-derivatives, it can be as simple as an extension of the combustor annulus. With can-type
combustors, the shape is quite complex, since it has to connect a round combustor can to an off-set nozzle
sector. This shape has been called a "goose neck", and is illustrated in Fig. 7.
Transition pieces carry gases at very high temperatures, about 2100 ºF (1150 ºC) in the machines of
1980's design, and approaching 2650 ºF (1450 ºC) in the latest "G" and "H" class machines, just
beginning to be installed. The inner surfaces of the transition piece may also "see" the primary
combustion zone, which subjects them to an additional radiant heat flux. Thus, they require heavy
cooling, and are usually coated with a ceramic on their inner, hot surface. In a through-flow can, they are
cooled by the secondary air. With reverse-flow cans, both the primary and secondary airflow passes
around them as it changes direction, facilitating their cooling.
Some of the latest designs use steam to cool the transition pieces, with the heat absorbed being utilised in
the steam cycle of a combined cycle plant. Fig. 8 shows a steam-cooled transition piece from the
Siemens-Westinghouse 501 G gas turbine.
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 M. A. Elmasri, 1990-2002
Chapter 4: GT Components - Combustors
The transition piece must provide adequate sealing where it meets the nozzle sector. This prevents
compressor discharge air from leaking directly into the turbine, by-passing the combustion section. This
requirement, plus the differential expansion and contraction as temperatures change, plus the need to
sustain differential vibrations between combustor liner and turbine nozzles render the transition piece
prone to cracking. As with all hot-section components, it poses many difficult design challenges.
Figure 8. Steam-cooled transition piece (501G, courtsey of Siemens-Westinghouse)
2.c. Silo Combustors:
This design has one or two enormous "cans", mounted on top of the machine or to its sides, rather than 816 smaller cans, mounted around its circumfrence. It has been used extensively by the European
manufacturers, Alstom and Siemens, but those manufacturers have recently been employing annular
combustors.
The silo combustor's large volume facilitates using a wide variety of fuels and provides long residence
time for good mixing and an even temperature profile at the turbine nozzles. It allows greater flexibility
in the design of the burner system, without the stringent volume constraints imposed on the smaller can or
annular combustors. Due to the large size, the structural shell may have to be lined with ceramic tiles to
lower its temperature and heat loss, and the large surface to be protected and cooled makes this design
bulky, heavy and expensive. With increasing pressure ratios and firing temperatures, the tendency has
been away from silo combustors.
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 M. A. Elmasri, 1990-2002
Chapter 4: GT Components - Combustors
Figure 9. Heavy-duty GT with silo combustor (GT11N2, courtesy of Alstom Power)
2.d. Annular Combustor:
Historically, this was primarily used in jet engines, and therefore aero-derivative machines, where its
main advantage is that it does not need a "bulge" in the machine's mid section, keeping frontal area small
and the shape aerodynamic. However, several recent designs of heavy-duty engines have employed this
concept, particularly in conjunction with a large number of small burners, with close circumferential
spacing, designed for low emissions. Other advantages are the reduced liner surface area to be cooled and
the absence of a complicated transition piece.
The annular combustor works best with high-pressure-ratio machines, where the high densities enable the
velocities to be reduced without needing a large flow cross-sectional area.
A disadvantage of some aeroderivative annular combustors is the need to disassemble the machine to
replace the annular liner.
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 M. A. Elmasri, 1990-2002
Chapter 4: GT Components - Combustors
Figure 10. Aeroderivative GT with annular combustor (LM1600, courtesy of GE)
Figure 11. Heavy-duty GT with annular combustor (GT13E2, courtesy of Alstom Power)
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 M. A. Elmasri, 1990-2002
Chapter 4: GT Components - Combustors
Figure 12. Heavy-duty GT with annular combustor (V84.3A, courtesy of Siemens-Westinghouse)
3. REHEAT OR "SEQUENTIAL" COMBUSTION
Fig. 13 shows a cross-section of the Alstom GT 24/26 reheat gas turbine and a photo of one being
assembled in the shop. The fundamental thermodynamic benefits of reheating the expanding turbine
gases have already been mentioned in the section on "Gas Turbine Cycle Thermodynamics".
The primary combustor faces design considerations similar to those for other gas turbines. The gases
leaving it expand through a single turbine stage, dropping in temperature by an estimated 550 ºF (300 ºC).
The total temperature of the hot gas leaving that stage and going into the reheat combustor is estimated to
be on the order of 1900 ºF (1040 ºC) and its oxygen content on the order of 13 %. This poses unique
problems in designing the second (reheat) combustor. The incoming hot gas cannot "cool" a liner, so
none is used, rather, the annular casing is protected by ceramic tiles, cooled by air. The burners are
probably also air-cooled, with the cooling air used to augment their oxygen supply. The reduced oxygen
content produces low flame peak temperatures, which is favourable in terms of reducing NOx production.
It is estimated that the primary combustor burns about ¾ of the total fuel, and the reheat combustor about
¼ of the total fuel.
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 M. A. Elmasri, 1990-2002
Chapter 4: GT Components - Combustors
Figure 13. Reheat gas turbine (GT 26, courtesy of Alstom Power)
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 M. A. Elmasri, 1990-2002
Chapter 4: GT Components - Combustors
4. POLLUTANT FORMATION & PREVENTION
A near-stoichiometric primary zone produces very high local temperatures, which allow the reaction
between N2 and O2 that forms oxides of nitrogen to occur. At normal temperatures, N2 and O2 do not
react, but do so only at elevated temeratures, above 2500º F (1370 º C). Nitrogen oxides, NO, NO2 and
N2O3, known collectively as NOx, have negative effects on human health, and contribute to "acid rain" by
forming nitric acid (HNO3) in the atmosphere. To reduce their formation, one needs to reduce the peak
temperatures in the combustion zone, as well as the time spent at, or near, peak temperature.
Before the 1990's NOx production was usually controlled by water or steam injection. Mixing steam or
water into the primary zone, reduces oxygen concentration and lowers peak flame temperature. To ensure
that the steam or water diluent is available in the combustion zone, it is either mixed with the fuel before
the fuel nozzles, or injected in close proximity to the fuel nozzles.
Water and steam injection have the undesirable side-effect of increasing other pollutants, notably CO and
unburnt hydrocarbons as they reduce NOx. They increase cost because they require demineralized water,
which is wasted with the flue gas and is expensive to produce. As a rule of thumb, water injection for
NOx control is usually at a flow rate roughly equal to the fuel flow; steam injection is usually at a rate of
about 1½ times the fuel flow. Thus, a 100 MW gas turbine using this form of NOx control would waste
about 200,000 tons of demineralized water per year.
GT power output increases with water and steam injection. The efficiency of the engine appears to
increase with steam injection, because by convention, the engine is not "charged" for the energy supplied
with the steam. The efficiency of the engine is reduced by water injection, since the energy supplied with
the water is low, and extra fuel is burnt to vapourise the injected water in the combustor.
If the GT is used in a combined cycle, any injected steam is diverted from the steam turbine, so power
loss from the steam turbine offsets power gain from the gas turbine. Water injection, on the other hand,
does not reduce power from the steam turbine, so a net gain in plant power is preserved, but efficiency
degradation is much worse than with steam injection.
Table 1 illustrates the effect of steam injection to lower NOx to 25 ppm in a plant employing a modern
aeroderivative gas turbine, burning gas in its "standard" combustor. The machine is also available, at a
premium, with dry low NOx combustion. The calculations include the effect of lowering the gas turbine's
TIT with injection to prevent it from overspeeding and to maintain its longevity.
Table 1. Example of the effect of steam or water injection for NOx reduction on thermodynamic
performance of the gas turbine and the combined cycle (GE LM2500+)
Gain in GT power output
with steam injection
+1.8 %
with water injection
+1.8 %
Increase in GT heat rate
-3.7 %
+7.5 %
Gain in combined cycle power output
-3.7 %
+3.3 %6
Increase in combined cycle heat rate
+7.2 %
+11.1 %
Another drawback of water injection is its tendency to produce vibration in the combustor, because the
water droplets boil explosively, creating accoustic waves and combustion instability. This can lead to
fatigue cracks or wear due to high-frequency rubbing between the combustor can and its locating
supports, or between the transition piece seals and the first-stage nozzle sectors.
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 M. A. Elmasri, 1990-2002
Chapter 4: GT Components - Combustors
Because of the above drawbacks, most manufacturers have developed dry low-NOx combustors, which
can achieve the environmental standards for air quality, at least when burning natural gas as a fuel,
without resorting to water or steam injection.
A typical way to reduce NOx formation without H2O injection is through lean combustion, which gives
lower flame temperatures. The fuel is first diluted with air, premixed with air in a proportion which is
either too fuel-rich or too fuel-lean to burn quickly on its own (flame propagation speed is fast in a
stoichiometric mixture, but slower in rich or lean mixtures). As the pre-mixed fuel/air reaches the
combustion zone, it is slowed down to create a stable flame front, and is mixed with secondary air if it
had originally been rich. The trick is to make sure that flame cannot flash back into the premixed,
air/fuel, by injecting the mixture at a velocity higher than
that at which the flame can propagate. It is relatively easy
to ensure this at the design-point, but less easy to ensure it
at varying loads, or with varying fuel composition. One
design is illustrated in Fig. 14.
Combustion
air
Spray
evaporation
Vortex
breakdown
Gas
Oil +
Water
Gas
Atomization
Gas injection
holes
Ignition
Flame
front
Figure 14. Low NOx burner (courtesy of Alstom Power)
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 M. A. Elmasri, 1990-2002
Chapter 4: GT Components - Combustors
In this design, with natural gas as a fuel, the premixed air/fuel is ejected in the form of a strong,
high-speed swirling vortex from the conical burner,
as illustrated in Fig. 14. It then burns as a "cool"
flame, as the vortex breaks down and the mixture
velocity drops at a stable flame front, standing in
space, away from the burner nozzles. Fig. 15 shows
a photo of the flame stabilized in space at some
distance from the burner.
When burning oil, however, some water must still be
mixed with the oil prior to its atomisation to lower
NOx production.
Fig. 16 shows an annular combustor utilising a large
number (29) of these burners. Since turndown is
usually limited in pre-mix, low-NOx burners, using a
large number of them allows some to be shut off at
low loads.
Figure 15. Natural gas flame from the low NOx
burner of Fig. 14 (courtesy of Alstom Power)
Figure 16. Annular combustor utilising many low NOx burners (courtesy of Alstom Power)
Most dry low-NOx burner designs can achieve NOx levels in the 9-25 ppm range when burning natural
gas. When burning oil, however, such levels may not be achievable, even with water injection.
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 M. A. Elmasri, 1990-2000
Chapter 5: GT Components - Turbines
GAS TURBINE COMPONENTS - TURBINES
Content Revised November, 2000
© Maher Elmasri 1990-2000
Figure 1a. A typical heavy-duty GT turbine (GT13E2, courtesy of Alstom Power)
The fluid dynamics of a turbine are very similar to those of a compressor, with the following notable
exceptions:
(i) Gases flowing through a turbine are moving in the direction of falling pressure. The pressure gradient
therefore "helps to push" the fluid in the viscous boundary layer, so it has no tendency to separate,
allowing turbine stages to operate with much higher turning angles, fluid accelerations, and pressure
ratios. Whereas a typical gas turbine with a 16:1 pressure ratio may have 18 compressor stages, its
turbine may only have 3, or 4 stages.
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 M. A. Elmasri, 1990-2000
Chapter 5: GT Components - Turbines
(ii) Because of the high turbine temperatures, the speed of sound, which is proportional to T1/2 , is high.
Turbines may therefore operate with subsonic flows at much higher tip and gas speeds. Nonetheless, they
are frequently designed with supersonic flow to reduce the number of stages.
(iii) The high temperature gas necessitates cooling the turbine components. This has a major impact on
the thermo-fluid dynamics.
Figure 1b. A typical heavy-duty GT first-stage turbine nozzle (Frame 7EA, courtesy of GE and
MassPower)
1. FUNDAMENTALS OF AN UNCOOLED TURBINE STAGE
1.a. Application of Conservation Laws of Momentum & Energy to an Uncooled Turbine
Stage:
Fig. 2 illustrates a turbine stage. The absolute gas velocity entering the rotor is c2 with a tangential
component of cy2 in the rotor's direction. The flow is turned in the rotor, exiting at an absolute velocity c3
with a tangential component cy3 against the rotor's direction. The rate of change of angular momentum of
air as it passes through the rotor must equal the torque τ,
τ = m r (cy2 + cy3) ..................... (1)
where r is the radius and m the gas mass flow rate. The power P developed by the rotor is the torque
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Chapter 5: GT Components - Turbines
times the angular velocity, ω, in radians/s,
P = m ω r (cy2 + cy3) .................. (2)
and the work done per unit mass of air, which must equal the drop in gas stagnation enthalpy h0, is
W = u (cy2 + cy3) = (h02 - h03) = (h01 - h03 ) ....... (3)
where u is the tangential speed of the rotor, u = ω r. The last equality is because the stator does no work,
so conservation of energy implies a constant stagnation enthalpy across it.
Stator
c2
Rotor
u=ωr
w2
w3
c3
β2
c2
α2
w3
w2
α3
β3
cX3
c3
u
cy2
cy3
Figure 2. Basic aerodynamics of a turbine stage
1.b. Relation between Enthalpy and Pressure Drop through an Uncooled Turbine Stage:
For a given pressure drop, the enthalpy drop depends on stage efficiency. If the stage is perfect, with
totally frictionless, isentropic flow, that enthalpy drop will be the maximum. All the pressure drop will be
utilized to decrease enthalpy and extract work from the gases. If the stage is totally inefficient, the flow
will be "throttled" and just drop to a lower stagnation pressure at constant enthalpy, producing no work.
If that continued until atmospheric pressure was reached, we would be left with hot gases, still containing
all their initial thermal energy, but having lost their pressure. They can no longer expand and provide
work. We may be able to recover work from their thermal energy using another heat engine, but that will
be less than what we could have recovered had we also utilized their pressure, not just their heat. In other
words, entropy would have been generated and exergy destroyed.
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Chapter 5: GT Components - Turbines
Ideal Stage:
For the perfectly efficient limit, if we assume a perfect gas, the stagnation (or total) enthalpy is directly
proportional to stagnation (total) temperature, and the isentropic pressure-temperature relationship can be
used to find the enthalpy and temperature ratios from the pressure ratio:
( h03,s / h01 ) = ( T03,s / T01 ) = ( P03 / P01 )(k-1)/k
........ (4)
The subscript s means "isentropic" in the above equation.
Real Stage:
The stage adiabatic efficiency (also called isentropic efficiency) is defined as the ratio (Actual Enthalpy
Drop/ Isentropic Enthalpy Drop) corresponding to a certain pressure ratio. Using the "total-to-total"
efficiency definition this is
h01 - h03
η = ----------- ...................... (5)
h01 - h03,s
and is illustrated on Fig. 3. The definition is a function of the assumed pressure ratio. For a very small
pressure ratio, the efficiency is called the Polytropic Efficiency, Є
Є = dh0/dh0,s ......................... (6)
h or T
h or T
01
p01
Real
Expansion
Isentropic
Expansion
h0
p0
03
p03
p0 – dp0
03s
h0 – dh0
h0 – dh0,s
s
s
Figure 3. Isentropic and real adiabatic expansion in an uncooled turbine
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 M. A. Elmasri, 1990-2000
Chapter 5: GT Components - Turbines
The small drop in enthalpy, temperature and pressure across the stage relate exit to inlet values by
h03,s = h01 – dh0,s
h03 = h01 – dh0
T03,s = T01 – dT0,s
T03 = T01 – dT0
p03 = p01 – dp0
.................... (7)
The binomial theorem reduces eq. 4 to
(dh0,s / h0) = (dT0,s / T0) = [ (k-1)/k ] (dp0 / p0) …….…(8)
and eq. 6 gives
(dh0 / h0) = (dT0 / T0) = [Є (k-1)/k ] (dp0 / p0) ………(9)
where the perfect gas assumption was used to linearly relate enthalpy and temperature.
Fig. 3 shows the T-s diagram for expansion of a perfect gas. Eq. 9 can be integrated for a finite, real
stage, or group of stages with a constant polytropic efficiency, to give
( h03 / h01 ) = ( T03 / T01 ) = ( P03 / P01 ) Є(k-1)/k
........ (10)
Efficiencies
Overall and exergy efficiencies of an expansion starting
at five times ambient temperature, with a polytropic
(small stage) efficiency of 0.9
0.98
0.97
0.96
0.95
0.94
0.93
0.92
0.91
0.9
Overall efficiency
Exergy efficiency
0
10
20
30
40
Pressure Ratio
Figure 4
For a given polytropic efficiency, Є, the adiabatic efficiency increases with increasing pressure ratio, as
seen in Fig. 4. Compounding many small stages, each with a polytropic efficiency of Є, results in a
higher overall efficiency, η, because the inefficiency of the upstream stages results in hotter gas reaching
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 M. A. Elmasri, 1990-2000
Chapter 5: GT Components - Turbines
the downstream ones. The work of expansion is higher with a hotter gas, so the downstream stages
produce more work than they would have produced, had the upstream stages been more efficient. On the
other hand, if one defines a second law efficiency as the ratio of (work output)/(decreased exergy of
gases), that exergetic efficiency falls with pressure ratio. The reason is that work dissipated into heat near
the beginning of expansion is at high temperature. That heat is therefore more convertible to work than
heat from dissipation at the lower temperature, near the end of expansion. Fig. 4 illustrates these results.
1.c. Relation between Stage Work and Blade Velocity & Angles - Uncooled Turbine:
For the case cx3=cx2, a constant axial velocity, eq. 3 can be combined with trigonometric relations from
Fig. 2 to give
(h01 - h03) = u2 { 1 + (tan β2 + tan α1) cx/u } ….. (11)
i.e. the stage work is proportional to the square of the peripheral speed and to the extent the flow is turned
by the blades.
The ratio (h01 - h03)/u2 is defined as the stage loading coefficient, ψ, and the ratio cx/u as the flow
coefficient, φ, so that in dimensionless form, eq. 11 is written as
ψ = 1 + φ (tan β2 + tan α1)
………… (11a)
Referring to the velocity triangles of Fig. 2, an observer moving with the rotor sees air coming in at the
relative velocity w2 and leaving at the relative velocity w3. In that relative coordinate frame, the rotor
appears stationary, doing no work. Thus the rotor-relative stagnation enthalpy, the sum of static enthalpy
and relative kinetic energy, is constant,
h2 + w22 /2 = h3 + w32 /2
....…………... (12)
so the static enthalpy drops across the rotor if the relative velocity, w, increases
h2 - h3 = (w32 - w22) /2
..…………......... (13)
Thus in the relative frame, the rotor behaves as a nozzle, building up relative kinetic energy and reducing
static enthalpy, as well as temperature and pressure. This build-up of relative velocity is needed to reduce
the absolute velocity (and total energy) left behind the rotor when the flow exits.
Both sets of blades act as nozzles in their respective coordinate frames, designed to accelerate the gases,
converting their pressure and enthalpy to kinetic energy. The kinetic energy developed in the nozzle is
absorbed in the rotor as work. The relative kinetic energy developed in the rotor serves to increase its
work absorption by leaving lower residual absolute velocity, enthalpy and pressure behind the rotor.
In the blade-relative frame, no work is done on the flow, and stagnation temperature and enthalpy are
constant. Relative stagnation pressure is constant across a frictionless, isentropic blade-row but falls a
few percent in a real case. In the absolute, stationary frame, the rotor absorbs the entire drop in stagnation
state, producing work.
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Chapter 5: GT Components - Turbines
1.d. Temperature "Seen" by a Blade Row:
A thermometer moving along at the same speed as a high speed gas will register its static temperature.
The reading of a stationary thermometer blocking the flow will closely approximate the gas stagnation
temperature. The difference between the two is the temperature rise corresponding to the kinetic energy
of the gas.
Due to frictional heating, an adiabatic (insulated and uncooled) surface, along which a high speed gas
flows, attains a temperature Taw, higher than that of the gas. This is called the Adiabatic Wall
Temperature or Recovery Temperature and is set by the balance between energy dissipation in the
viscous boundary layer, which heats it up, and the transfer of the resulting heat into the cooler main
stream flowing by.
If the abovementioned surface were cooled to a lower temperature, Tw, the heat transfer rate from the gas
to the surface would be proportional to the difference (Taw - Tw) where Taw is the adiabatic wall
temperature and Tw the actual wall temperature.
The Recovery Factor is defined as the difference between adiabatic wall and gas static temperature
divided by that between stagnation and static:
Recovery Factor = (Taw - Tg)/(T0 - Tg)
This is a function of Prandtl Number, a fluid property. For most gases this is slightly below unity, but it
is common to approximate it as unity, which implies Taw = T0, and provides conservative estimates for
surface cooling calculations. With the assumption of Recovery Factor = 1, for heat transfer and cooling
calculations, the surface "sees" gas at its stagnation temperature.
The nozzle sees the gas at the inlet stagnation temperature, T01. The fact that the gas accelerates through
the nozzle and the static temperature T2 drops far below T1 (typically by 300º-450º F or 170º-250º C),
does not reduce the heat transfer, since the stagnation temperature is constant throughout the acceleration
and T02 = T01.
The rotor, on the other hand, sees the Rotor-Relative Stagnation Temperature T0,rr which is the sum of
static temperature plus kinetic energy relative to the rotor blades. Although the relative velocity
accelerates from w2 to w3 in the rotor, the relative stagnation temperature remains constant. This is
because in the rotor-relative frame no work transfer is observed and the rotor "looks like a nozzle", with a
constant relative stagnation enthalpy.
The rotor-relative stagnation enthalpy is therefore,
h0,rr = h2 + w22 /2
....................... (14)
= h02 - (c22 - w22 )/2
............... (15)
and the temperature seen by the rotor, T0,rr is below that seen by the stator by
T02 – T0,rr = (c22 - w22 )/2 Cp ............ (16)
This difference can be several hundred degrees. Increasing it can significantly lower the cooling load on
the rotor surfaces. It is a function of the velocity triangles and aerodynamic design. For example, an old
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Chapter 5: GT Components - Turbines
GE Frame 5 gas turbine, which has a pressure ratio of ten with only two turbine stages, is estimated to
have a first-stage nozzle exit velocity c2 of almost 2,600 ft/s (790 m/s), which corresponds to a Mach
Number of about 1.3. The blade-relative inlet speed, w2, is estimated at about 1390 ft/s (425 m/s). This
corresponds to a difference between absolute and rotor-relative stagnation temperatures of about 320ºF
(180ºC). Thus, the machine is operated with an absolute rotor inlet stagnation temperature of 1755ºF at
full load, without needing to cool the rotor, which only sees about 1435ºF. For a Frame 7F, first-stage
difference between absolute and rotor-relative stagnation temperatures is estimated at about 300ºF
(167ºC).
1.e. Degree of Reaction and Effect on Turbine Cooling:
Similar to the compressor, Reaction is defined for the turbine as the static enthalpy drop through the rotor
divided by the total stage enthalpy drop:
R = (h2 - h3) / h01 - h03)
If c3 = c1, then
......................... (17)
R = (h2 - h3) / (h1 - h3)
An Impulse Turbine has 0 % reaction. The nozzle accelerates the flow, which impinges on the turbine
"bucket". The rotor-passage has a constant flow cross section, it just turns the flow, with a constant
relative velocity, and constant static pressure, temperature and enthalpy. Such a design has the advantage
of no leakage across the rotor blade tips since there is no pressure difference across the rotor. Moreover,
it has the advantage of minimizing T0,rr. Its primary disadvantage is poor aerodynamic efficiency.
Accomplishing the entire static pressure drop in the nozzle implies high nozzle exit velocity, c2,
frequently supersonic, which reduces nozzle efficiency due to development of shock waves. Also, the
high angle of turning in the bucket at constant mainstream pressure can create viscous boundary layer
separation on the concave wall, further increasing losses.
50% reaction implies equal apportionment of static enthalpy drop between rotor and stator. It also
implies symmetric velocity triangles if cx is constant. The flow velocities are lower than for an impulse
stage. Turning the flow in the rotor happens under a favourable pressure gradient, reducing tendancy for
boundary layer separation. The aerodynamic efficiency is therefore higher than for an impulse stage.
However, the difference between T02 and T0,rr is less than for an impulse stage, requiring more rotor
cooling. Also, for a given blade speed, u, the pressure ratio is lower than that for an impulse stage, which
means more stages, hence even more turbine cooling.
Fig. 5 shows the velocity triangles for an impulse stage and a 50% reaction stage, both with the same
blade speed, u, and both with axial exit, c3. Typical numerical values are also listed.
100 % reaction would have the entire drop in static state in the rotor, the stators being used only to
redirect the flow at constant velocity and pressure. This design has nothing to recommend it.
Low reaction raises stage loadings and reduces the number of stages required. The cooling penalties are
reduced since there are fewer blade stages, end-walls, and wheel spaces to cool. The large decrease
between T02 and T0,rr serves to further reduce the need for cooling. Turbine stages are expensive, so
capital cost is reduced by high loadings. The main drawback of low-reaction designs is lower stage
aerodynamic efficiencies. The compromise between those factors normally leads to optimum reactions of
about 20-30% in cooled turbines.
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 M. A. Elmasri, 1990-2000
Chapter 5: GT Components - Turbines
β2
α2
w2 = w3
β2 = β3
w3
c2
Impulse stage with axial exit and ψ = 2
α3 = 0
β3
w2
w2 = c3
α2 = β3
β2 = 0
50% reaction stage with axial exit and ψ = 1
α3 = 0
c2 = w3
w2
α2
cy3 = 0
u
cy2 = 2u
c3
Typical values:
u=1280 ft/s (390 m/s)
cx = 780 ft/s (240 m/s)
c2 = 2680 ft/s (820 m/s)
w2 = 1500 ft/s (460 m/s)
W = Δh0 = 131 BTU/lb (305 kJ/kg)
ΔT0 = 470 ºF (260 ºC)
PR = 2.25
T02 – T0,rr = 350 ºF (195 ºC)
β3
u
cy2 = u
Typical values:
u=1280 ft/s (390 m/s)
cx = 780 ft/s (240 m/s)
c2 = 1500 ft/s (460 m/s)
W = Δh0 = 66 BTU/lb (153 kJ/kg)
ΔT0 = 235 ºF (130 ºC)
PR = 1.5
T02 – T0,rr = 120 ºF (65 ºC)
c3
cy3 = 0
Figure 5. Examples of pure impulse and 50% reaction velocity triangles
1.f. Three-Dimensional Effects:
All the arguments given above are for a 2-dimensional, pitch-line analysis. Since the peripheral speed, u,
increases with radius, the velocity triangles and aerodynamics vary along the radius.
If the energy extraction from the gas were to be constant at all radii, the loading parameter, ψ, would have
to vary inversely with the square of the radius. The blades would have to be twisted to increase flow
turning at the hub and decrease it at the tip. Degree of reaction would be smallest at the hub and largest at
the tip. This is called a "Free-Vortex" design since fluid absolute tangential velocities are inversely
proportional to radius.
One problem with free-vortex designs is that the last turbine stage, which would have a small hub/tip
radius ratio, would need considerable blade twist. Assuming that were mechanically feasible, the
aerodynamic design suffers, since a reasonable pitch-line loading could correspond to an unreasonablyhigh loading at the hub radius, where the degree of reaction could become zero or negative. Negative
reaction implies diffusion in the rotor passage, with attendant boundary layer separation problems. Those
factors force the designer to reduce stage loadings towards the rear of the turbine. They also limit the
maximum efficient overall pressure ratio for a single turbine spool to about 20, although a value of 30 has
been used in the Alstom GT 24/26. Higher pressure ratios on a single spool will result in either very short
blades at its inlet, with high leakage and end-wall friction losses, or very long blades at its exit, with
excessive blade twist or highly non-uniform radial loading.
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Chapter 5: GT Components - Turbines
2. FUNDAMENTALS OF TURBINE COOLING
2.a. The Need for Turbine Cooling:
The first-stage blades of a modern, heavy-duty gas turbine have a pitch speed of about 1300 ft/s (400
m/s), faster than the muzzle velocity from a shotgun or pistol. They are subjected to a centrifugal
acceleration of about 16,000 g's (16,000 times gravity), so a blade which weighs 5 lbs (2.25 kg) is
subjected to a centrifugal force of about 40 tons. In addition, the gas exerts a bending force on the blade
of over half a ton. All this while the blade sees a temperature of over 2000 ºF (1100 ºC).
Conventionally cast or forged nickel-based alloys used for turbine blading allow nozzles to operate at
about 1550 ºF (845 ºC), and permit rotor blade temperatures of about 1400 ºF (760 ºC). Advanced
material processing has been pushing allowable rotor blade temperatures up by providing higher strength
at elevated temperatures. Directional Solidification and Single Crystal Blade technologies can each
contribute an increment of about 40 - 60 ºF (20 - 35 ºC) to rotor blade temperature, so blade metal
temperatures in the 1500 ºF (815 ºC) range are now attainable. Fig. 6 shows a directionally-solidified
blade casting, before it is machined. The molten metal is cooled in a controlled way as it solidifies,
producing crystals which are aligned with the highly-stressed radial direction, which improve strength in
the direction where it is most needed.
Figure 6. DS blade casting (courtesy of Siemens-Westinghouse)
Maintaining nozzle metal temperatures in the 1600 ºF (870 ºC) range and rotor blade metal temperatures
in the 1500 ºF (815 ºC) range, requires blade cooling, since in a modern heavy duty GT, the first-stage
nozzles see gas at about 2500 ºF (1370 ºC), and the first-stage rotor blades see gas at about 2100 ºF (1150
ºC). Thus, the cooling system needs to maintain the blade metals below the adiabatic wall temperature
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Chapter 5: GT Components - Turbines
(the temperature of the gas that they "see"), by 600-900 ºF (330-500 ºC). To attain this, the cooling
system needs to be augmented by thermal barrier coatings on the metal surfaces. Most turbines are aircooled. Although much research has been done on water-cooling and on steam-cooling, the latter is just
beginning to be applied in commercial designs. Thus, the discussion below will focus mostly on aircooling.
2.b. Cooling Air Circuit :
The cooling air is extracted from the compressor. Air for cooling the first stage nozzles must be from
compressor discharge to ensure adequate pressure for flow through the nozzle's internal cooling passages
and subsequent discharge into the gas stream. The gas pressure around the first rotor, except at the
stagnation point on the front "nose" of the blade, is well below compressor discharge pressure, since the
gas static pressure drops in the first nozzle by a factor of 1.5 to 2. Nevertheless, it is common to extract
all the cooling air for the first-stage nozzle and rotor from the compressor discharge, to ensure adequate
cooling-air pressure.
The temperature at compressor discharge is about 750 ºF (400 ºC) for a pressure ratio of 15; and about
1000 ºF (540 ºC) for a pressure ratio of 30. To economise in usage of cooling air, some designs pass the
extracted cooling air through a heat exchanger to cool it, typically down to about 400 ºF (200 ºC). The
heat removed from the cooling air is utilised in the bottoming cycle, typically to make extra steam and/or
heat feedwater. In some designs, part of that heat is utilised to preheat the GT's fuel, thereby reducing its
consumption. Even when that heat is recovered, pre-cooling the cooling air introduces thermodynamic
losses, since exergy is destroyed by moving heat from a hotter to a cooler medium. Figure 7 illustrates
the cooling circuit of a heavy-duty gas turbine. The cooled cooling air is re-introduced into the rotor,
where it flows through axial passages, then radially outwards through wheelspaces, thence into holes in
the blade platforms and through the blades.
Figure 7. Cooling circuit of a heavy-duty GT (501F, courtesy of Siemens-Westinghouse)
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Chapter 5: GT Components - Turbines
Subsequent rotor stages at are lower pressure. However, to simplify the design and ensure adequate
cooling air pressure, it is common to extract most, if not all, of the rotor cooling air from compressor
discharge (or one or two compressor stages before it).
The second and third stators are usually cooled with air extracted from intermediate bleeds, at lower
pressure and temperature.
Apart from cooling the nozzles and blades, the turbine rotor and/or disks, bearings and end walls are all
cooled as well.
Some designs use an external booster compressor to raise the pressure of the extracted cooling air, but
such a system introduces substantial, additional parasitic losses, and is thus of dubious efficacy.
2.c. Thermodynamic Penalties due to Cooling:
Air cooling the turbine introduces three types of losses:
(i) Gas Quenching: Due to the combination of heat absorption by the coolant from the gas through the
blade walls and the subsequent introduction of spent coolant into the hot gas path.
(ii) Main Gas Stagnation Pressure Loss: Arising from the introduction and mixing of a low momentum
gas (the coolant) into a high-speed stream (the main gas). This is sometimes expressed as an increase in
"profile loss" - a measure of stagnation pressure loss in flow around an airfoil - with coolant injection.
(iii) Coolant Throttling: Representing the exergy destruction due to the difference in stagnation pressure
between the point of coolant extraction in the compressor and the point at which it re-joins the turbine
work-producing gas stream.
The tabulation below illustrates the relative magnitude of exergy destruction by each, for an example of a
hypothetical turbine, based on parameters estimated for a three-stage turbine of 1990's design.
Table 1. Exergy destruction, expressed as a percentage of the GT fuel LHV heat input, by various
mechanisms in a cooled turbine
Thermal
Mainstream
degradation pressure loss
Coolant
throttling
Cooling
totals
Adiabatic
loss
Total
First Stage
1.01
0.559
0.385
1.954
0.441
2.395
Second Stage
0.372
0.366
0.605
1.343
0.444
1.787
Third Stage
0.073
0.043
0.051
0.167
0.454
0.621
Totals
1.455
0.968
1.041
3.464
1.339
4.803
It is worthwhile to note that the cooling loss in the first stage is about five times the adiabatic loss, which
occurs due to aerodynamic friction, regardless of whether the stage were cooled or not. Had the turbine
been adiabatic (uncooled), about 3.5 percent of the cycle's exergy supply would not have been destroyed,
thus the GT efficiency would have been about 40%, instead of about 36.5%.
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 M. A. Elmasri, 1990-2000
Chapter 5: GT Components - Turbines
2.d. Thermodynamic Modelling of an Air-Cooled Turbine Stage:
Total (Stagnation) Specific Enthalpy
All Pressures are
Total (Stagnation)
PA
PB=PA-∆Pm,u
A
∆hm,u
B
B’
Expansion at
Adiabatic Efficiency
∆hw
Expansion at
"Cooled Efficiency"
PC=PD+∆Pm,d
PD
C
∆hm,d
C’
D
Specific Entropy
Figure 8. Model of expansion through an air-cooled turbine stage
One approach to analyzing the thermodynamics of a cooled stage is to separate the enthalpy and
stagnation pressure drops due to cooling from those due to work-extraction. Figs. 8 and 9 illustrate such a
model.
The hot gas enters the stator at State A. Due to quenching by the stator coolant, the stagnation specific
enthalpy drops by Δhm,u to hB, and a loss in stagnation pressure occurs because of entraining the coolant
into the fast-flowing gas. Thus, the mass flow entering the rotor is augmented by the "non-chargeable"
stator cooling air, but its thermodynamic state is degraded from State A to State B. One then assumes that
the gases expand through the rotor with the same, adiabatic efficiency as they would have expanded
without cooling, from State B to State C. State C, however, is at a stagnation pressure higher than the
true stage back-pressure, by the loss occasioned by mixing the rotor coolant with the fast-flowing gas.
Finally, quenching the hot gas by the rotor cooling air further degrades the specific enthalpy by Δhm,d ,
from hC down to hD , before the gas enters the next stage, augmented in mass flow but degraded in
temperature.
It is interesting to note that a cooled turbine expansion path on T-s coordinates can show decreasing
specific entropy, since State D can have a lower specific entropy than State A. Lowering the mean
specific entropy due to introducing cool air can outweigh the entropy generated by all irreversibilities.
This is not a violation of the second law, since total entropy is increased by the entropy generated, but the
specific entropy, per unit mass of turbine-path gases, may be reduced.
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 M. A. Elmasri, 1990-2000
Chapter 5: GT Components - Turbines
Enthalpy
hA
Δhm,u
hB
c12/2
hA,s
Δhrr=(c22- w22)/2
Δhw
hrr
c22/2
(w32- c32)/2
hB,s
hC
Δhm,d
hD
Stator
Total Enthalpy
Static Enthalpy
Rotor
hD,s
c3 /2
2
Distance along flow
Figure 9. Total and static states for the model of expansion through an air-cooled turbine stage
2.e. The Cooling Effectiveness Curve:
Minimizing the amount of cooling air is essential to preserve the benefits of higher firing temperatures.
Progress in this direction has been made by the development, during the last two decades, of DS and
Single Crystal blades, as well as thermal barrier coatings, which are thin layers of ceramic bonded to the
blade's exterior surface. Along with these developments, much work has been done to optimise the
intricate channels and holes in the blade through which the coolant passes. The earliest designs were
simple convective channels within the blading. Those have evolved into complicated inserts with small
jets of coolant impinging upon the inner surfaces of the blade skin, or flowing through ribbed or finned
passages to enhance heat transfer, as illustrated in Figure 10. Advanced designs also use film-cooling,
where the cooling air is ejected from thin holes to flow along the surface of the blade, shielding it from
the hot gas, as may be seen in Fig. 1b. The ultimate film-cooling concept, transpiration cooling, with a
porous blade from which the coolant emanates, has been the subject of much R&D but is not in practical
use.
The capability of a blade-cooling design is customarily defined by the "Cooling Effectiveness",
Φ = (Tg – Tw) / (Tg – Tc) ………….. (18)
where the subscripts are:
g: gas adiabatic wall temperature (approximately equal to wall-relative stagnation temperature)
w: wall
c: coolant at supply to blade
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Chapter 5: GT Components - Turbines
Cooling air
Blade
Vane
Fins
Pins
Fins
Ribs
Fir
tree
foot
Cooling air
Ribs
Ribs
Figure 10. Cooling passages in a first-stage (courtesy of Alstom Power)
The cooling effectiveness increases with coolant flow rate for a given blade design. It is therefore
common to correlate it with Blowing Ratio, BH, the ratio of coolant heat capacity to main gas heat
capacity. This correlation typically takes the form shown below.
where
and
BH = α [Φ / (Φ∞ - Φ ] n
…………………………. (19)
BH = (m Cp)coolant / (m Cp)gas
…… …………………. (20)
α is a coefficient, bearing a proportionality to the area ratio between the cooled blade wall and the
gas flow cross-section, and to the external heat transfer coefficient
n is an exponent, typically between 0.8 and 1.25
Φ∞ is the effectiveness that would result from an "infinite" coolant blowing rate, typically 1 with
film cooling and about 0.85 with convective cooling
Fig. 11 illustrates typical cooling effectiveness curves. 100 % effectiveness implies that the blade is
maintained at coolant temperature. 0% effectiveness implies that the blade is as hot as the gas total
temperature. The curves shown are estimated for a first-stage blade row of 1990's design. For the filmcooling curve, the parameters used in eq. 19 are α = 0.045, Φ∞ = 1, and n = 0.9. For the convective
cooling curve, the parameters used are α = 0.0576, Φ∞ = 0.85, and n = 1.25. These curves assume no
TBC (thermal barrier coating). Ceramic Thermal Barrier Coatings (TBC) are currently used on combustor
liners and on blades of advanced engines. Adding this barrier improves the metal-substrate cooling
effectiveness but degrades the coating-surface value. In Fig. 11, the curve labelled "Film/TBC" assumes
a 0.012" (0.3 mm) coating of Yittria-stabilized Zirconia. The effectiveness it reflects defines the
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Chapter 5: GT Components - Turbines
temperature of the metal substrate, not the coating temperature, which would be hotter. For a first-stage
blade row, such a coating reduces the required cooling air flow by about 30%, for a given temperature of
the metal substrate. The application of TBC to all stators and rotors in a modern engine can reduce total
coolant requirements by about 15% and improve GT cycle heat rate by about 5%.
1
Film
0.9
Film/TBC
Cooling Effectiveness
0.8
Convective
0.7
0.6
0.5
0.4
0.3
0.2
0.1
0
0
5
10
15
Blowing Ratio = (m Cp)coolant / (m Cp)gas
Figure 11. Cooling effectiveness curves
Figure 12. First-stage nozzles, showing film-cooling holes near the trailing edge.
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 M. A. Elmasri, 1990-2008
Chapter 6: GT Installation & Environment
GAS TURBINE INSTALLATION & ENVIRONMENT
Content Revised October, 2002, Updated September, 2008
© Maher Elmasri 1990-2008
In this chapter we shall discuss various aspects of the gas turbine installation and environment, and their
impact on performance. The discussions are all from the perspective of a fixed gas turbine, and how its
operating parameters are affected by its environment, such as altitude, ambient temperature and humidity,
the inlet and outlet pressure losses, and the type of fuel. Methods of conditioning the inlet air to improve
performance are discussed in Chapter 13.
6.1
GAS TURBINE INLET SYSTEM
Figure 1. Components in the air/gas path of the gas turbine (courtesy of G&H Acoustics)
Modern gas turbines need inlet air filtration. To prevent excessive noise levels, the inlet and exhaust
ducts also need silencers, and the exhaust needs a stack to direct flue gases upwards and discharge them
far enough from ground level. These components, illustrated in Figure 1, introduce pressure losses in the
air as it enters, and in the exhaust gases as they leave the gas turbine. In this section, we shall discuss the
inlet and exhaust equipment and the thermodynamic impact of these pressure losses on performance.
Additional equipment which may be installed in the inlet of a gas turbine to boost performance, such as
inlet air evaporative coolers, foggers, refrigeration coils or supercharging fans, is discussed in Chapter 13.
6.1.1
INLET AIR FILTRATION
Older gas turbines, especially those used for peaking applications, in which they operated for a limited
number of hours per year, were sometimes installed without an inlet air filter, only a screen to prevent the
ingestion of large bugs or birds. While this may seem surprising to some, it is worthwhile remembering
that aircraft jet engines operate satisfactorily without inlet air filtration. Naturally, much of the operating
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Chapter 6: GT Installation & Environment
life of a jet engine is spent at high altitude, where the air is relatively free of the larger dust or sand
particles found near sea level.
The air in metropolitan areas in the USA contains total suspended particles in the range of 40-160
micrograms per cubic meter, which corresponds to roughly 0.035-0.14 ppm by weight. The concentration
of suspended particles in the air is much higher near heavy industry and in parts of the world which are
more dusty or polluted, and can be much higher, reaching hundreds of ppm during desert sandstorms.
Even with a relatively clean 0.1 ppm by weight, a 180-MW class gas turbine without a filter would ingest
about 1¼ tons per year of solid particles. This would erode the smooth coatings on the compressor blades
as well as pose the dangers of plugging film cooling holes and causing erosion damage to the turbine
blade coatings as well. Suspended particles which contain sodium or potassium, such as in salty air near
the ocean, or dust from soils rich in these minerals, such as former seabeds, are particularly harmful since
they also cause hot corrosion in the turbine. Hence the need for air filtration for modern gas turbines.
Figure 2. Replaceable panel filters (courtesy of G&H Acoustics)
Figure 3. Self-cleaning cartridge filters (photo by the author, courtesy of Masspower)
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Chapter 6: GT Installation & Environment
There are two types of inlet air filters, those with replaceable panels which hold the dirt until they are
replaced, illustrated in Figure 2, and the self-cleaning cartridge type, on which the dust layer forms until it
is blown off, illustrated in Figure 3. The characteristics, such as pressure drop and air-cleaning efficiency,
are generally similar for both types. Both types can remove virtually 100% of the larger particles (above
10 µm), and both can attain pressure drops of less than 25 mb (1” H2O) when new and clean. The main
difference is that the self-cleaning type can maintain a relatively low pressure drop over a longer period of
time, and therefore requires less frequent replacement, especially in dusty environments which would
rapidly clog the replaceable panel type. On the other hand, the self-cleaning type is more complex, since
it requires a compressed air system with valves that periodically open to create a puff of compressed air
that reverses the airflow through the cartridge elements, thereby dislodging the accumulated dust. These
puffs are implemented in a cyclic pattern, going sequentially through banks of cartridges, so that only a
small percentage of the cartridges are puffed at any particular instant, preventing any tangible shock to the
gas turbine’s inlet air flow rate. Sceptics have argued that the dislodged dirt stays in the vicinity of the
filter, and gets re-entrained in the air and re-deposited on the filter elements, but tests have shown the
efficacy of this system. This arises from the observation that the fine particles stick together on the filter
surface, forming a cake. When they are puffed away, they are in the form of a larger, heavier
agglomeration, too heavy to become fluidized and re-entrained by the low velocity airflow. These
heavier particles are either blown away by the wind or cleaned off the structure.
Self-cleaning cartridge filter are generally preferable in adverse environments which would otherwise
necessitate frequent and costly replacement if panel filters are used. Panel filters are simpler and less
expensive if the environment is clean.
The filter elements are pleated and usually made of paper or a woven synthetic fiber. For replaceable
panel filters, typical design face velocities based on the projected area of the pleated filter are 2-3 m/s (610 ft/s). Since the surface area along the pleats is 5-8 times as large as the projected area, the actual
velocity of the air as it goes through the filtration surface is only 1/5th to 1/8th of the face velocity.
Roughly 0.4 m2 of filter face area are needed for each kg/s of GT airflow (2 ft2 for each lb/s of airflow).
Thus, a modern gas turbine in the 180-MW class needs a total filter face area of about 200 m2 (2000 ft2).
For self-cleaning cartridge filters, the pleats are usually shallower and thus the superficial velocities based
on the cylindrical surface area of the filter are much lower, about 0.5 m/s (1.6 ft/s). Thus, one cartridge
with a diameter of ≈0.3 m (12”) and a length of ≈0.6 m (24”) will usually be installed per 0.45 kg/s (1
lb/s) of airflow.
6.1.2
INLET STRUCTURE
Figure 3 shows a typical inlet air housing used with cartridge type self cleaning filters. It consists of
several plenum chambers, each fed through numerous cartridge filters. The plenums all feed into the inlet
duct behind them, which leads downwards into the gas turbine inlet bellmouth.
In the photographs of Figure 3, about 300 cartridge filters are fitted into each of three plenums, for a total
of about 900 cartridges for the entire housing, which feeds a gas turbine with a nominal airflow of about
300 kg/s (about 660 lb/s). The cleaning puffs are implemented to groups of eight cartridges at a time, and
the entire cleaning process, implemented once per day, takes about half an hour.
Figure 4 shows an inlet housing for a filter of the replaceable panel type. The filter panels are vertical and
protected by weather hoods.
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Chapter 6: GT Installation & Environment
Figure 4. Typical inlet housing with panel filters & silencers (courtesy of G&H Acoustics)
In addition to holding the inlet filters, the intake structure and its ductwork also houses the silencers, as
well as the anti-icing system, if used. When a gas turbine is operated at an ambient temperature between
–2 °C (30 °F) and +4 °C (39 °F), there is the possibility of ice forming on the walls of the inlet ducts and
the engine’s bellmouth. If chunks break away and enter the engine, they may cause damage. Below –2
°C (30 °F), there is too little vapour humidity in the air to form ice within the gas turbine inlet structure,
and the ambient humidity is already in the solid phase, as particles of snow or ice, will either be caught in
the inlet filters or be too small to cause any damage. Above +4 °C (39 °F) the air is too warm to form ice
within the gas turbine inlet structure. The danger is present mostly when the ambient is humid and just
above the freezing temperature, since the air acceleration around bends in the inlet structure, or as it
enters the bellmouth, causes a few degrees drop in temperature, which allows humidity carried in as
vapour or liquid to form ice on the walls of the ducts or the gas turbine inlet. Icing problems are most
likely to be encountered in plants right along the ocean (or gas turbines on platforms), as well as
installations where shifting winds blow a cooling tower plume towards the gas turbine inlet.
Figure 5. Inlet housing with anti-icing hot air distribution upstream of the filters (courtesy of G&H Acoustics)
To prevent ice formation, the inlet air is heated when conditions warrant it. Heating may be
accomplished by bleeding hot, compressed air from the gas turbine’s compressor and mixing it with the
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Chapter 6: GT Installation & Environment
incoming ambient air. The hot air may be bled from the compressor’s discharge, or from one of the
intermediate bleeds used during startup and to extract turbine cooling air. Since the required temperature
rise is usually less than 5 °C, less than 2% of the engine’s air would be needed from an intermediate bleed
at 250 °C, and even less from the compressor discharge. The hot air bled must be uniformly mixed into
the inlet air, usually via perforated pipes within the intake, downstream of the inlet filters. In some
installations it is mixed upstream of the filters, using a multiplicity of perforated pipes to distribute it, as
shown in Figure 5.
In installations where inlet air chilling is used, the finned tube coil used for the chilled water may also be
used with hot water to heat the inlet air. The hot water may come from one of the heat recovery boiler’s
economisers. In district heating or cogeneration plants, the warm water returning from the process is a
convenient source of “free” heat for the inlet air.
6.1.3
INLET SILENCER
The compressor of a gas turbine creates a loud whine, principally at the blade passing frequency. For
instance, a 3600 RPM, 60-Hz machine with 40 blades on the first compressor stage will have a blade
passing frequency of 2.4 kHz. Since other stages have a different number of blades, and the number of
vanes on the IGV ring usually differs from the number of first stage blades, other various frequencies of
noise are created by blade-wake interactions, mostly in the range ½ to 5 kHz, a range acutely felt by the
human ear as an annoying whine.
To minimise loud noise from emanating from the compressor to the surroundings, the plenum or duct
leading to it is fitted with sound-absorbing walls and silencers, such as shown in Figure 6. This consists
of a series of compliant panels through which the air flows. The panel walls are made of a soft material,
typically a dense synthetic fiber matrix enclosed in thin perforated stainless steel sheets. These panels
absorb the acoustic pressure fluctuations, and attenuate the high intensity sound waves propagating
through the passages. The pressure drop incurred by the silencers is usually quite small relative to that
incurred by the inlet air filters.
Figure 6. Inlet air silencer panels (courtesy of G&H Acoustics)
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Chapter 6: GT Installation & Environment
6.1.4
EFFECT OF INLET PRESSURE LOSS ON GT PERFORMANCE
The inlet pressure drop comprises the loss in the filter as well as in the inlet silencer and ductwork.
Additional pressure losses may also arise from optional equipment, such as finned tube coils for inlet air
chilling or heating.
6.1.4.1 Effect of Inlet Pressure Drop on GT Airflow and Cycle T-s Diagram
Figure 7 illustrates the impact of the inlet pressure drop on the gas turbine cycle’s T-s diagram, using
ideal gas laws and a vastly exaggerated 20% inlet pressure drop for clarity of the diagram. The solid lines
1-2-3-4 show the cycle without an inlet loss, and the dashed lines 1’-2’-3’-4’ show it with an inlet loss,
for the same gas turbine.
The inlet to the engine is throttled from state 1 to a lower pressure, but the same temperature, state 1’.
Because the inlet pressure drop reduces the density of compressor inlet air, in linear proportion to
pressure, the air mass flow rate entering the engine is also reduced in linear proportion to pressure. Thus,
assuming the firing temperature is held constant by the control system (T3’=T3), the fact that less mass
flow is being forced into the fixed-area turbine nozzles results in a proportionate reduction in turbine inlet
pressure, and hence in compressor delivery pressure. Thus, p2’ is below p2 by the same percentage that
p1’ is below p1, both reductions being the same as the percentage reduction in air mass flow rate; so the
pressure ratio across the compressor is unchanged; so the temperature ratio across the compressor is also
unchanged (thus T2’=T2). The pressure ratio across the turbine, rp,T, however, falls, since p4’ is the same
as p4 but p3’ is below p3 by the same percentage as the reduction in air mass flow rate. Thus,
δrp ,T
rp ,T
=
δp3
p3
=
δp 2
p2
=
δp1
p1
=
δmair
.................... (1)
mair
Effect of Inlet Pressure Drop on Cycle T-s diagram (exaggerated DP/P of 20% )
3
1200
3'
1100
1000
900
Temperature, C
800
700
600
500
400
2
300
4
2'
4'
200
100
0
1
-100
-0.1
1'
0
0.1
0.2
0.3
0.4
0.5
0.6
0.7
0.8
0.9
1
1.1
Specific Entropy, kJ/kg-K
Figure 7. Thermodynamic effects of inlet air pressure drop illustrated on T-s diagram
The atmosphere at sea level has an absolute pressure of 1013 mb or 407" H2O, thus, a typical inlet
pressure drop of 10 mb (4" H2O) causes a reduction of about 1% in the air mass flow rate, as well as in
the compressor inlet pressure p1, as well as in the compressor delivery pressure p2, and also in the turbine
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 M. A. Elmasri, 1990-2008
Chapter 6: GT Installation & Environment
inlet pressure p3. The turbine exit pressure, p4, is unaffected, so the turbine pressure ratio, rp,T, falls by 1%
as well.
6.1.4.2 Effect of Inlet Pressure Drop on GT Exhaust Temperature
The reduction in turbine pressure ratio raises the exhaust temperature for a fixed T3. To estimate the
magnitude of the exhaust temperature rise, one may write the adiabatic expansion equation as
T4 =
T3
 p3
 xT

p 4 

=
T3
rp ,T
xT
..................................... (2)
where rp,T is the pressure ratio across the turbine (rp,T = p3/p4). For an adiabatic expansion with a
polytropic efficiency ηT, the turbine exponent xT is given by
xT = ηT
(γ − 1)
γ
.................................................. (3)
where γ is the specific heat ratio of the combustion gases. Perturbing equation (2) by taking the partial
derivative of T4 with respect to the turbine pressure ratio rp,T, then simplifying, gives the relationship
between a small change in turbine pressure ratio and the corresponding change in exhaust temperature for
a fixed turbine inlet temperature:
δT4
T4
= − xT
δrp ,T
rp ,T
......................................... (4)
Since γ for combustion gases is about 1.33, for an 87% polytropic efficiency, xT is about 0.22. So that a
1% reduction in inlet pressure, hence in turbine pressure ratio, corresponds to an exhaust temperature
increase of roughly 0.22%. Since typically T4 is of the order of magnitude of 530 °C (of the order 800 K),
δT4 would be of the order 0.0022 x 800 or about 1.75 °C (3.15 °F). In real gas turbines, the relationship
between turbine pressure ratio and temperature ratio is more complicated than equation (2), due to turbine
cooling, and the controls do not keep T3 perfectly constant. In most practical heavy duty gas turbines, a
1% throttling of inlet pressure results in an increase of exhaust temperature of between 1 and 2 °C (1.8
and 3.6 °F).
6.1.4.3 Effect of Inlet Pressure Drop on GT Output and Efficiency
The inlet pressure drop reduces power output by two mechanisms:
(a) The reduction in engine airflow, which, on its own, results in a proportionate reduction in power
output. Since a 1% inlet pressure drop reduces airflow by 1%, it thus reduces power output by 1% as
well.
(b) The reduction in turbine pressure ratio, with the compressor pressure ratio remaining unchanged. It
was shown in §6.1.4.2 above, that the inlet pressure drop raises T4 by δT4 without affecting T1, T2, or T3.
Thus, turbine work is reduced in proportion to δT4, whereas cycle net work is roughly proportional to the
difference between turbine work and compressor work, i.e.
δW
Wnet
≅
C p ,T δT4
C p ,T (T3 − T4 ) − C p ,C (T2 − T1 )
......................................... (5)
where Cp,T and Cp,C are the mean specific heats for the turbine and compressor, respectively, about 1.03
and 1.23 kJ/kg-K, respectively. Substituting for δT4 from equation (4) and expressing the temperature
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 M. A. Elmasri, 1990-2008
Chapter 6: GT Installation & Environment
ratios across the turbine and compressor through their pressure ratios, i.e. equation (2) for the turbine and
its equivalent for the compressor gives
δW
Wnet
where
xT = ηT
≅
(rp ,T
(γ − 1)
xT
 δr

− xT  p ,T
C p ,T
r
p ,T 

x
rp ,T T
x
− 1)C p ,T −
(rp ,C C − 1)C p ,C
........................ (6)
ϑ
is the exponent for an adiabatic turbine expansion, about 0.22 ........ (6a)
γ
(γ − 1)
xC =
is the exponent for an adiabatic compression, about 0.31 ................... (6b)
γηC
rp,T and rp,C are the pressure ratios across the turbine and compressor, respectively, typically
about 16 for a modern, heavy duty gas turbine ........................................................... (6c)
ϑ is the cycle temperature ratio, T3/T1, typically about 5.5 ....................................... (6d)
With the typical values shown above, and for a 1% inlet pressure drop, equation (6) gives an estimate for
the reduction in net output attributed to the reduced turbine pressure ratio of about 0.5%.
Thus, for a 1% inlet pressure drop, the total estimate for the reduction in net output is about 1.5% , 1%
arising from the reduced air mass flow rate, and 0.5% arising from the reduced turbine pressure ratio.
The estimated efficiency drop would be only 0.5% (percent of the efficiency, not percentage points),
arising from the reduced turbine pressure ratio only, since the reduction in air mass flow rate is
accompanied by a proportionate reduction in fuel flow rate, because the combustor temperature rise is
unaffected, since T2 and T3 are the same.
The above is a simplified, but reasonable, analysis. In real gas turbines, the analysis is more complicated
than equation (6), due to turbine cooling, and to the fact that the controls do not keep T3 perfectly
constant. In most practical heavy duty gas turbines, a 1% reduction in inlet pressure results in a loss of
power in the range of 1.25%-1.75%, and an increase in heat rate in the range of 0.25%-0.75%.
6.1.4.4 Effect of Inlet Pressure Drop on CC Output and Efficiency
The effect of inlet pressure loss on the entire combined cycle is less severe than on its gas turbine. Since
the inlet loss reduces gas turbine efficiency, it increases exhaust energy as a proportion of fuel input.
Furthermore, the increased exhaust temperature and reduced exhaust mass flow rate both improve the
efficiency of the heat recovery boiler. Thus, although steam turbine output declines with GT inlet
pressure drop, its decline as a percentage is very much smaller than the decline in the gas turbine’s output,
so the whole plant’s output declines less than the gas turbine’s output on a percentage basis. In typical
combined cycles, if the inlet pressure drop reduces gas turbine power output by ∆%, it reduces combined
cycle net output by only about ¾ ∆%.
The effect on heat rate is even weaker. If a certain GT inlet pressure drop increases gas turbine heat rate
by δ%, it increases combined cycle net heat rate by only about ¼ δ%.
6.2
GAS TURBINE EXHAUST SYSTEM
The velocity of the gases leaving the last stage of a gas turbine is high, frequently exceeding 250 m/s (820
ft/s). At 250 m/s, the “velocity head” or “dynamic pressure” of the gases , ½ ρ V2, is about 130 mb or 50
inches of water, too high to be wasted. An exhaust diffuser, usually supplied or designed by the gas
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 M. A. Elmasri, 1990-2008
Chapter 6: GT Installation & Environment
turbine manufacturer, is used to smoothly slow these gases down and recover as much of their exit kinetic
energy as possible, converting it to pressure and reducing the static pressure behind the last turbine stage.
The flue gas velocity after such a diffuser is about one-third or one-fourth of that leaving the last stage of
the gas turbine, i.e. in the range 50-75 m/s (160-250 ft/s). At 50 m/s, the dynamic pressure of the gases is
just about 5 mb (2 “H2O), and this is typically wasted, since additional diffusers are rarely efficient due to
the space constraints on their design.
In open cycle gas turbine installations, the stack follows the diffuser, with an intermediate section of high
temperature duct as dictated by the site geometry. In installations with a heat recovery boiler, a large
transition duct is used to further slow down the gases, by a factor of 3-4, before they enter the boiler’s
main duct, which contains the tubes. Typically, due to space limitations, the rate of divergence necessary
for such a transition duct is so great that it cannot be an efficient diffuser. When the gases have to be
turned, a smooth bend, frequently fitted with turning vanes, is used to minimise pressure losses.
Exhaust ducts and high-temperature stacks must withstand the full GT exhaust temperature. They are
usually made of double-wall construction, with a strong, gas-tight carbon steel shell and an internal, thin,
corrosion resistant stainless-steel liner. Insulation between the two layers keeps the external carbon steel
casing at a moderate temperature. Figure 8 shows the inside of a duct, with the frame that supports the
internal hot liner.
Figure 8. Exhaust ductwork showing inner liner supports & insulation space (courtesy of G&H Acoustics)
6.2.1
BY-PASS DAMPERS
Some installations with heat recovery have by-pass (or “diverter”) dampers and stacks. These allow the
gas turbine exhaust to be diverted to the by-pass stack before it reaches the heat recovery boiler, allowing
the gas turbine to operate without the boiler. In some cases, the diverter damper position is used to
modulate the proportion of the exhaust that enters the boiler, hence to reduce boiler heat input and steam
production as needed.
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 M. A. Elmasri, 1990-2008
Chapter 6: GT Installation & Environment
Figure 9. By-pass (or “Diverter”) damper (courtesy of G&H Acoustics)
Because of their large size and their ability to handle hot gases, exhaust gas diverter dampers are
expensive. Since they divert the exhaust gases to a high-temperature stack, this too is expensive
compared to a low-temperature stack downstream of the heat recovery boiler. Adding a diverter damper
and a by-pass stack to a 180-MW-class gas turbine installation can add about $4-5 million in equipment
costs, and about $8-10 million in total costs, after considering the additional ductwork, foundations, and
labour, as well as the expanded footprint of the unit. In addition to the capital costs, large diverter
dampers are seldom leak-free. If only 0.5% of the flue gases escape, the value of the lost recoverable heat
from a 180-MW-class gas turbine is roughly $450,000 per year (at $12/million BTU of recoverable heat).
Another complicating factor arises if the power plant uses SCR or other pollution controls within the
HRSG, since these will be by-passed when the damper is open. This could add further constraints on the
permitting process and may restrict the number of operating hours per year when the by-pass can be used.
Thus, use of by-pass dampers and stacks should be specified with caution, and only after considering the
alternatives. One alternative solution for cogeneration plants with periods of low steam demand is the
installation of a high-pressure dump condenser that condenses unwanted steam during periods when the
power requirements compel the gas turbine to be run at a load that would produce more steam in the
HRSG than the process requires.
It is unfeasible to perform HRSG repairs with the gas turbine operating and exhausting through the bypass stack. Concerns about exhaust gas leakage, as well as heating up of the HRSG interior by radiation
and conduction, make it impermissible to send plant personnel into the HRSG.
6.2.2
STACKS & SILENCERS
Double-wall insulated stacks must be used for high gas temperatures, such as for simple cycle
installations without a HRSG, or by-pass stacks upstream of the HRSG. A typical construction is shown
in Figure 10, with stainless steel liner separated by insulation from a carbon steel exterior structure.
Single-wall, uninsulated stacks are sometimes used after heat recovery boilers with low stack
temperatures. In these cases, exterior insulation and cladding must be installed in areas within reach of
personnel, as a safety measure. With low stack temperatures in cold environments it may be necessary to
insulate the entire stack, especially if the ambient is windy, to prevent condensation within the stack.
Most stacks contain a silencer consisting of a multiplicity of compliant baffles, similar to the intake
silencer described in §6.1.3, except that it is made of temperature resistant materials. Stacks following a
heat recovery boiler may, in some cases, omit the silencer, since the boiler’s insulation lined duct, and its
numerous rows of densely packed finned tubes may adequately impede the transmission of turbine noise
to the exterior.
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 M. A. Elmasri, 1990-2008
Chapter 6: GT Installation & Environment
Figure 10. Interior of a double-wall insulated stack (left) and stack silencer baffles (right) (courtesy of G&H Acoustics)
6.2.3
STACK PRESSURE DROP
The stack imposes a pressure drop due to fluid friction within the stack and due to the dynamic pressure
of the leaving loss, which is wasted in the atmosphere. On the other hand, the column of hot gas in a tall
stack is lighter than the surrounding air, which tends to create a natural draft, countering the friction
pressure drop and leaving loss.
6.2.3.1 Stack Leaving Loss
Combined cycle stacks are usually sized to provide a flue gas velocity of about 20 m/s (65 ft/s). This
guarantees that the gases are ejected at a sufficient speed to augment their buoyancy, ensuring that they
rise as they mix with the ambient air. Typical flue gases after a heat recovery boiler in a combined cycle
are at roughly 100 °C, and have a density of about 0.92 kg/m3. At 20 m/s exit velocity, their leaving
kinetic energy corresponds to a pressure loss of ½ ρ V2 = ½*0.92*202 = 184 Pa (≈1.8 mb ≈0.72 “H2O).
Stacks for simple cycle installations, or by-pass stacks, are usually sized for higher velocities than
combined cycle stacks. Due to the low gas density, a sizing velocity on the order of 20 m/s (65 ft/s)
would result in an expensive stack with an excessive diameter, so a reasonable sizing velocity is on the
order of 45 m/s (150 ft/s), which results in a mass flux similar to that for a combined cycle stack. Typical
exhaust from a heavy duty gas turbine is at roughly 575 °C (1067 °F), with a gas density of about 0.4
kg/m3. At 45 m/s gas dynamic pressure is about ½ ρ V2 = ½*0.4*452 = 405 Pa (≈4 mb ≈1.6 “H2O).
6.2.3.2 Stack Friction Pressure Drop
At 20 m/s, a stack for a modern 180-MW-class gas turbine combined cycle, would have a diameter of
about 6 m (20 ft). If the stack height is 60 m (200 ft), its length/diameter ratio would be ten. Assuming a
mean roughness of ε/D=10-4 would yield a friction pressure drop of only ≈0.25 mb (0.1 “H2O). At 45
m/s, a stack for the direct exhaust of a modern 180-MW-class gas turbine would also be about 6 m (20 ft)
in diameter. If it is 60 m (200 ft) tall, its friction pressure drop would be about 0.5 mb (0.2 “H2O).
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Chapter 6: GT Installation & Environment
Stacks for plants with smaller gas turbines would have a smaller diameter, but if the required height is the
same as for a large plant, for air permit considerations, the length/diameter ratio, which sets friction
pressure drop, is higher for the smaller plant. In any event, friction pressure drop is generally much
smaller than the dynamic pressure leaving loss.
6.2.3.3 Stack Buoyancy Draft
The density of ambient air at 15 °C is 1.22 kg/m3. Thus, density difference between a combined cycle
stack at 100 °C and the ambient at 15 °C is ∆ρ=1.22-0.92=0.3 kg/m3. A stack 60 m (200 ft) tall would
thus create a buoyancy draft of ∆ρ g h = 0.3 * 9.81 * 60 = 176 Pa (1.74 mb or 0.7 “H2O). Density
difference between GT exhaust at 575 °C and normal ambient is ∆ρ=1.22-0.4=0.82 kg/m3. A 60 m (200
ft) stack directly following a heavy duty gas turbine would thus create a buoyancy draft of ∆ρ g h = 0.82
* 9.81 * 60 = 483 Pa (4.8 mb or 1.9 “H2O).
Thus, a typical stack for a large combined cycle would not create any net pressure loss at all, since the
buoyancy would just about cancel the leaving loss and friction. The same holds true for many simple
cycle or by-pass stacks, which have higher leaving loss but also greater buoyancy.
6.2.4
EFFECT OF EXHAUST PRESSURE LOSS ON GT PERFORMANCE
The exhaust system pressure drop is usually about 5-10 mb (2-4 “H2O) for an installation without heat
recovery, and about 15-30 mb (6-12 “H2O) for installations with heat recovery. Other than the pressure
drop within the heat recovery boiler, including catalysts if present, most of the losses are dynamic
pressure lost in bends, sudden expansions, and the exit to the atmosphere.
6.2.4.1 Effect of Exhaust Pressure Drop on GT Cycle Performance
Unlike inlet loss, exhaust loss does not reduce airflow. It does, however, reduce turbine pressure ratio for
a given compressor ratio. The result is a higher exhaust temperature and a loss in both power and
efficiency, both comparable in magnitude to the efficiency loss resulting from an equivalent inlet pressure
drop.
Effect of Exhaust Pressure Drop on Cycle T-s diagram (exaggerated DP/P of 20% )
3'≅3
1200
1100
1000
900
Temperature, C
800
700
600
4'
4
500
2'≅2
400
300
200
100
0
1'=1
-100
-0.1
0
0.1
0.2
0.3
0.4
0.5
0.6
0.7
0.8
0.9
1
1.1
Specific Entropy, kJ/kg-K
Figure 11. Thermodynamic effects of exhaust pressure drop illustrated on T-s diagram
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 M. A. Elmasri, 1990-2008
Chapter 6: GT Installation & Environment
Figure 11 illustrates the impact of the exhaust pressure drop on the gas turbine cycle’s T-s diagram, using
ideal gas laws and a vastly exaggerated 20% exhaust pressure drop for clarity of the diagram. The solid
lines 1-2-3-4 show the cycle without an exhaust loss, and the dashed lines 1’-2’-3’-4’ show it with an
exhaust loss, for the same gas turbine. Any propagation of the disturbance in exhaust pressure to
upstream points in the cycle is negligibly small. Thus, the only effect is the drop in the pressure ratio
across the turbine, rp,T, whilst the pressure ratio across the compressor remains the same. The result is an
increase in exhaust temperature, per equation (4), and a loss in net power output, per equation (6).
Because there is no change in airflow or fuel flow, the percentage loss in efficiency is the same as the
percentage loss in power output.
Thus, following the analysis in §6.1.4.2, a 1% exhaust pressure drop with a constant firing temperature
results in an increase of exhaust temperature of between 1 and 2 °C (1.8 and 3.6 °F). Following the
analysis in §6.1.4.3, it results in a loss of power in the range of 0.25%-0.75%, and an increase in heat rate
of 0.25%-0.75%.
6.2.4.2 Effect of Exhaust Pressure Drop on CC Output and Efficiency
The reduction in GT output and efficiency occasioned by the exhaust pressure drop results in an increase
in exhaust energy. This energy increase is in the form of higher temperature, so the percentage increase
in exhaust exergy is even greater than the percentage increase in exhaust energy. This improvement in
grade of heat is reflected in an improvement in heat recovery efficiency. Therefore the decline in
combined cycle performance is less severe than the decline in gas turbine performance. With typical
assumptions, the percentage decline in combined cycle output and heat rate is about one-quarter of the
decline in gas turbine output and heat rate.
The discussions given above are based on the assumption of a constant gas turbine cycle firing
temperature, T3. As discussed later, T3 may be controlled by measuring T4 and adjusting for turbine
pressure ratio, rp,T. If rp,T is indeed measured, through measuring both p4 and p3 (or p2, assuming that
combustor percentage pressure drop is constant), then T3 can be indeed maintained constant. If, on the
other hand, only p3 is measured, by measuring p2 and assuming that combustor percentage pressure drop
is constant, and p4 is assumed constant for the installation, then small changes in p4 during operation will
not cause T4 to change, since they do not affect the measured p2. Instead, they will result in changes in
the firing temperature, T3, at constant exhaust temperature, with different results.
6.3
EFFECT OF AMBIENT TEMPERATURE
In this section, we use the term “Ambient Temperature” to mean the gas turbine compressor inlet
temperature. These are essentially the same when there is no cooling of the gas turbine inlet air,
discussed in Chapter 13. In the absence of inlet cooling, compressor inlet air is slightly warmer than
ambient, due to heat conduction and radiation from the hot gas turbine engine to its inlet filter housing.
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Chapter 6: GT Installation & Environment
6.3.1
EFFECT OF COMPRESSOR INLET TEMPERATURE ON GT PERFORMANCE
Effect of Increased Ambient Temperature on Cycle T-s diagram
3 3'
1200
1100
1000
900
Temperature, C
800
700
600
500
2'
2
400
4 4'
300
200
100
0
1
-100
-0.1
0
1'
0.1
0.2
0.3
0.4
0.5
0.6
0.7
0.8
0.9
1
1.1
Specific Entropy, kJ/kg-K
Figure 12. Thermodynamic effects of ambient temperature on T-s diagram
Figure 12 illustrates the impact of ambient temperature on the gas turbine cycle’s T-s diagram, using ideal
gas laws. The solid lines 1-2-3-4 show the cycle at the normal 15 °C (59 °F) ambient, and the dashed
lines 1’-2’-3’-4’ show it at an extremely hot ambient of 50 °C (122 °F), for the same gas turbine. Since
air density varies inversely with absolute temperature, the air mass flow rate entering a machine of given
size and rotational speed is reduced on a hot day. This reduction in air mass flow rate reduces the cycle’s
pressure ratio, since less pressure is needed to force the reduced mass flow through the fixed turbine
nozzles, causing p3 to fall. A lower p3 reduces turbine pressure ratio, rp,T , so that exhaust temperature T4
rises, for a fixed firing temperature, T3.
6.3.1.1 Effect of Ambient Temperature on GT Airflow
If the compressor inlet were operating with low Mach numbers, the incompressible flow approximation
would apply, meaning that volumetric flow rate would be constant for a fixed geometry and rotational
speed. This would imply an air mass flow rate proportional to 1/T1, i.e. inversely proportional to the
absolute ambient temperature. If, on the other hand, the compressor inlet were choked, with sonic or
supersonic velocities, the mass flow for fixed geometry and speed would be proportional to 1/√T1, i.e.
inversely proportional to the square root of absolute ambient temperature. Most older gas turbines have
subsonic speeds at the compressor inlet, and the newer designs run with supersonic velocities at the first
stage blade tips and subsonic velocities at the hubs. Thus, although compressibility effects are important,
the airflow is not fully choked, causing mass flow rate to vary in proportion to 1/T1n. With a fixed RPM
and fixed inlet guide vanes, the exponent n is typically between 0.8 and 1. If the guide vanes are
adjustable, or the compressor speed variable, such as in certain aeroderivative gas turbines, the exponent n
could be outside this range.
Assuming that mair ~ 1/T1n gives
δmair
mair
= −n
δT1
T1
.................... (7)
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Chapter 6: GT Installation & Environment
Thus, if the ambient temperature increased from the ISO value of 15 °C (59 °F) to 25 °C (77 °F), the
relative change in absolute temperature would be δT1/T1 = 10/288.15 ≅ 3.5 %, so if n were 1, say, the air
mass flow rate would decrease by about 3.5% per 10 °C increase in compressor inlet temperature and
would increase by about 3.5% per 10 °C decrease in compressor inlet temperature, i.e.
δmair
mair
≈ -3.5% per 10 °C δT1 ........ (7a)
6.3.1.2 Effect of Ambient Temperature on GT Compressor Delivery Temperature
With the turbine behaving as a choked nozzle, p3 is proportional to mair, and assuming that the percentage
combustor pressure drop is unchanged, p2 is also proportional to mair, and so are the compressor and
turbine pressure ratios, rp,C and rp,T, respectively. Thus
δrp , C
rp , C
=
δ r p ,T
r p ,T
=
δp 2
p2
=
δp 3
p3
=
δmair
mair
= −n
δT1
T1
.................... (8)
Since
T2 = T1rp ,C
xC
.......................................... (9)
Thus, the change in T2 is
δT2 =
∂T
∂T2
δT1 + 2 δrp ,C
∂rp ,C
∂T1
................. (10)
which, after manipulation with equation (8) gives
δT2 = rp ,C x δT1 − nxC rp ,C x δT1 ............ (11)
C
C
where xC is the compression exponent, given in equation (6b). The first term reflects the direct effect of
T1 on T2, amplified by the nominal temperature ratio across the compressor, and the second term reflects
the effect of the changed compressor pressure ratio on T2.
As an example, with a compressor pressure ratio rp,C of 16, an exponent xC of 0.31, and n=1, say, one gets
δT2=1.63 δT1; i.e. a 10 ° increase in ambient results in a 16.3 ° in compressor discharge temperature. In
machines which directly use some of the compressor discharge air as the cooling medium of the first
turbine stage, this stage will run hotter on a warm day, because it will receive hotter coolant. To estimate
the effect on blade metal temperature, we use the cooling effectiveness equation,
Φ=
which gives
Tg − Tb
Tg − Tc
∂Tb
=Φ
∂Tc
......................................... (12)
............................. (13)
For example, for a heavy duty gas turbine with Tg=1200 °C (2192 °F), Tb=850 °C (1562 °F), and Tc=400
°C (752 °F), the effectiveness Φ is about 0.44, so the blade metal temperature increases by 44% of the
increase in coolant temperature. Thus, a 10 ° rise in compressor inlet temperature T1 causes an increase
of about 7 ° in blade metal temperature (0.44*16.3).
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Chapter 6: GT Installation & Environment
Since blade material lifetime can be halved if it runs 15 °C hotter than normal, an increase of 20 °C (36
°F) in ambient temperature can halve the blade lifetime, since it increases blade metal temperature by
about 14 °C. Thus, a gas turbine that does not precool the cooling air taken from compressor discharge,
and is installed in a tropical environment, without compressor inlet air cooling, would have about half the
hot stage blade lifetime of a similar model operating in an ISO environment.
In some gas turbines the cooling air for the first turbine stage is pre-cooled in a heat exchanger. The heat
sink for the heat exchanger may be ambient air or boiler feedwater or steam. If the heat sink is ambient
air, then the cooling air reaching the turbine blades will be warmer on a hot day, but its temperature
increase with ambient will be less than for designs where the cooling air is used directly, without precooling. If the cooling medium is boiler water or steam, then the cooling air temperature of reaching the
turbine blades will not be sensitive to ambient conditions.
6.3.1.3 Effect of Ambient Temperature on GT Exhaust Temperature
For a fixed firing temperature, T3, the change in exhaust temperature, T4, can be found from equation (4),
after substituting for δrp,T from equation (8) and manipulating:
δT4 = nxT
ϑ
rp ,T
xT
δT1
.................................................. (14)
where xT and ϑ are given by equations (6a) and (6d), respectively. For example, with n=1, xT=0.22,
θ=5.5, and rP,T=16, one gets:
δT4=0.66 δT1 ................................. (14a)
i.e. the change in exhaust temperature is about two-thirds the change in ambient temperature for a typical
heavy duty gas turbine.
6.3.1.4 Effect of Ambient Temperature on GT Power Output
Power output is affected by the change in air mass flow rate, and, additionally, by the change in overall
cycle temperature ratio, θ=T3/T1, occasioned by change in T1 at constant T3. Increasing T1 reduces
airflow, which reduces power; additionally, it reduces θ, which reduces specific power, further reducing
gas turbine power output. Thus, the percentage decline in power exceeds the percentage decline in air
mass flow rate.
To quantify the relative change in net power, we can write, approximately,
δW
Wnet
δm   − C p ,T δT4 − C p ,C (δT2 − δT1 ) 
≅  +

 m   C p ,T (T3 − T4 ) − C p ,C (T2 − T1 ) 
............................. (15)
where the first term represents the effect of mass flow rate reduction, and the second term the effect of the
drop in specific power. After some manipulation, this can be expressed in terms of the change in ambient
temperature as:


 ϑ 
 − C p ,C (rp ,C xC − nxC rp ,C xC − 1) 

− nC p ,T xT 
 r xT 
 δT1
δW 
 p ,T 
≅ − n +

Wnet 

1 
 T1
xC

−
−
−
C
1
C
(
r
1
)
ϑ
p ,T
p ,C
p ,C


 r xT 
p ,T




6-16
........... (16)
 M. A. Elmasri, 1990-2008
Chapter 6: GT Installation & Environment
For example, with n=1, xC=0.31, xT=0.22, θ=5.5, rP,C=rP,T=16, Cp,T=1.23, Cp,C=1.03, one gets for a typical
heavy duty gas turbine that (δW/Wnet) ≈ {-1-0.86}(δT1/T1); and since each 10 °C change in ambient
represents a 3.5% change in T1, this means that
δW
Wnet
≈ -6.5% per 10 °C of δT1 ........ (16a)
6.3.1.5 Effect of Ambient Temperature on GT Heat Rate
An increase in T2 causes an equivalent reduction in combustor temperature rise, and thereby in fuel heat
input per unit of airflow. Thus, the change in fuel heat input, q, is the change in airflow plus the change
in fuel input per unit airflow:
δq
q
≅
δmair
mair


− C p ,C δT2
+

 0.5(C p ,T + C p ,C )(T3 − T2 ) 
.............. (17)
which, with equations (7) and (11), after simplification, gives






2C p ,C
δq 
 1 − nxC  δT1
≅ − n −

(C p ,T + C p ,C )  ϑ − 1  T1
q 


xC



 rp , C

............ (18)
Where the first term represents the reduction in fuel input occasioned by the reduction in airflow at
warmer ambients, and the second the reduction occasioned by the hotter compressor discharge at warmer
ambients. For our typical heavy duty gas turbine numerical example, with n=1, xC=0.31, θ=5.5, rP,C=16,
Cp,T=1.23, Cp,C=1.03, one gets (δq/q) ≈ {-1-0.47}(δT1/T1); and since each 10 °C change in ambient
represents a 3.5% change in T1, this means that
δq
q
≈ -5.1% per 10 °C of δT1 ........ (18a)
and the percentage change in efficiency η (or negative change in heat rate, HR) is the difference between
the percentage change in power output and the percentage change in fuel input, i.e.,
δη
δ ( HR) δWnet δq
=−
=
−
Wnet
q
( HR)
η
.............. (19)
which can be evaluated through equations (16) and (18). For our specific numerical example of the
typical heavy duty gas turbines, the values given in equations (16a) and (18a) yield,
δη
δ ( HR)
=−
≈ -1.4% per 10 °C of δT1 ........ (19a)
( HR)
η
so for our typical heavy duty gas turbine numerical example, each 10 °C increase in ambient causes a
1.4% increase in heat rate, and a 1.4% drop in efficiency (not 1.4 percentage points!).
6.3.2
TYPICAL TEMPERATURE CORRECTION CURVES FOR HEAVY DUTY GT'S
Manufacturers supply “correction curves” that show how the gas turbine performance is affected by
compressor inlet temperature. These usually show correction factors on the y-axis for each of the four
6-17
 M. A. Elmasri, 1990-2008
Chapter 6: GT Installation & Environment
cardinal performance quantities: power output, exhaust mass flow rate, exhaust temperature and heat rate
(or efficiency), versus compressor inlet temperature on the x-axis. Multiplying the value of each
performance quantity at nominal ISO temperature (15 °C or 59 °F) by its correction factor gives its value
at any compressor inlet temperature.
120
-15
-10
-5
5
10
15
20
25
30
35 C
Power
115
Correction Factor, %
0
Heat Rate
110
105
100
Exhaust Flow
95
90
85
Lines show Thermoflow model
Symbols show vendor data
80
75
0
1040
20
30
40
50
60
70
80
90
100
110
120
Line shows Thermoflow model
Symbols show vendor data
1030
Exhaust Temperature [°F]
10
1020
1010
1000
990
980
970
960
950
940
0
10
20
30
40
50
60
70
80
90
100
110
120
Compressor Inlet Temperature [°F]
Figure 13. Typical correction curves for a heavy duty gas turbine
Since power, exhaust flow, and heat rate are all absolute quantities with an obvious datum of zero, they
are unambiguously determined by the correction curves, which apply to any set of units. When exhaust
temperature is described through a correction factor, however, some ambiguity can arise, since it is not
obvious whether the correction factor should be multiplied by absolute temperature (in Kelvin or
°Rankine), which would be the logical, measure, or whether it should be multiplied by the temperature in
°C or °F, which, apart from its irrationality, would require different correction curves for different
systems of units. Therefore it is common for manufacturers to supply three correction curves, on one
graph, for power, heat rate, and mass flow rate, and supply a second graph showing exhaust temperature
explicitly, or as a difference from its value at ISO ambient conditions. Figure 13 shows an example.
6-18
 M. A. Elmasri, 1990-2008
Chapter 6: GT Installation & Environment
A few manufacturers supply correction factors for exhaust temperature, creating some ambiguity about
whether they should be multiplied by absolute exhaust temperature, which makes the curve independent
of units, or by temperature in customary scales such as °C or °F, which requires different curves for
different units. An example is shown in Figure 14.
5
14
23
Effect of Ambient Temperature - Typical Heavy Duty GT
32
41
50
59 F
68
77
86
95
104
115
Parameter, % of ISO value
110
105
Power
100
Airflow
T exhaust (C)
95
Heat Rate
90
85
80
-15
-10
-5
0
5
10
15
20
25
30
35
40
Ambient Temperature, C
Figure 14. Typical correction curves for a heavy duty gas turbine
The curves shown on Figures 13 and 14 are for different models of heavy duty gas turbines, made by
different manufacturers . They are consistent with the values given in the foregoing analysis. The
approximate effects per 10 °C (18 °F) increase of compressor inlet temperature on a typical gas turbine
are summarised in Table 1:
Table 1. Effects of changing compressor inlet temperature on a typical gas turbine
Parameter
Approximate change per 10 °C (18 °F)
increase of T1
Analytical
Estimate
Air & exhaust mass flow rate
- 4%
Equation (7a)
Power output
- 7%
Equation (16a)
+ 1½%
Equation (19a)
+7 °C (+13 °F)
Equation (14a)
Heat rate
Exhaust temperature
Some models, such as the one represented by the curves of Fig. 14, will show a change in slope in the
power curve at very cold ambients. This may be due to a maximum power limit, based on mechanical
strength of the shaft, couplings, or thrust bearings; or on limiting generator or transformer capacity.
Alternatively, it may be due to a maximum pressure ratio limit on the compressor, to avoid potential
compressor surge. If the change in slope of the power curve is accompanied by a change in slope in
exhaust temperature, then the manufacturer is limiting power output and pressure ratio via reduction of
T3. If the change in slope of the power curve is accompanied by a change in slope in the airflow curve,
then the manufacturer is limiting power output and pressure ratio via closure of the IGV’s.
6-19
 M. A. Elmasri, 1990-2008
Chapter 6: GT Installation & Environment
6.3.3
TYPICAL TEMPERATURE CORRECTION CURVES FOR AERODERIVATIVE GT'S
Figure 15 shows three shaft layouts used in aeroderivative gas turbines. Layout (a) is the oldest, simplest
method of converting a jet engine to generate shaft power, in which the pressurised hot gas which would
normally by ejected through the propulsion nozzle is passed through a new power turbine instead. In this
layout, the compressor and the turbine stages driving it are frequently termed the “gas generator”, since
they just produce hot, pressurised gas. The turbine stages driving the compressor are termed the “gas
generator turbine”, abbreviated GGT in Fig. 15, and the power turbine which drives the load is labelled
PT in Fig. 15. If the engine is used for generator drive, the PT rotates at a fixed speed, set by the
generator frequency regardless of load or ambient, but the GGT rotates at its own speed, since it is not
mechanically coupled to the PT. Examples of aeroderivatives employing this configuration include the
Rolls-Royce Avon and the GE LM2500. It is also a common layout for two-shaft, heavy duty gas
turbines used for both mechanical drive and generator drive applications, such as the Alstom GT10, and
the Solar Mars.
1
2
4
3
3P
G
GGT
Compressor
PT
(a) 2-shaft with free power turbine
1
2
3
4
3L
3P
G
LPC
HPC
HPT
LPT
(b) 3-shaft with free power turbine
1
2
3
PT
4
3L
G
LPC
HPC
HPT
LPT
(c) 2-shaft with common LP/power turbine
Figure 15. Shaft layouts for aeroderivative gas turbines
6-20
 M. A. Elmasri, 1990-2008
Chapter 6: GT Installation & Environment
Jet engines with higher pressure ratios have used a two-shaft gas generator, and the power generation
derivations of them have employed Layout (b) in Figure 15. The PT is not mechanically coupled to either
of the gas generator spools, so each of these spools rotates at its own speed. The LP spool is on a shaft
that passes through a hollow shaft connecting the HP spool turbine and compressor. Ball/roller bearings
locate the concentric shafts and allow them to rotate at different speeds. Examples of gas turbines
employing this configuration include the Rolls-Royce RB211, the Pratt & Whitney FT4 and FT8, and the
GE LM1600 and LM5000.
Layout (c) is a more recent configuration, essentially a modification to Layout (b) that integrates the
power turbine with the LPT, such that the enlarged LPT drives both the LPC and the load on the same
shaft. The load may be driven from the engine’s hot-end (as illustrated in Fig. 15) or from its cold-end.
This ties the LPC speed to the generator’s, set by the grid frequency, so it can no longer float with load
and ambient, improving stability. This configuration reduces the potential for overspeeding the power
turbine upon a loss of load, since the drag from the LPC will help to slow it down upon fuel shutoff.
With configurations (a) and (b), even after fuel shutoff triggered by a loss of load, the power turbine
continues to accelerate, since the inertia of the gas generator spool(s) keep them rotating and pumping air
through the unloaded power turbine. The GE LM6000 uses Layout (c), and the Rolls-Royce Trent is
similar in that the LPC and LPT are on the load shaft, but it has two additional free spools, not one.
For multi-shaft engines, the free spools have to be in equilibrium, their turbine work being exactly
balanced by their compressor work and bearing losses. Because the free spools can float in rotational
speed to find an equilibrium speed which depends on ambient and load, the correction curves are more
complex than for single-shaft machines. To illustrate the principals, we shall focus our discussions first
on the simpler configuration shown in Layout (a).
In configurations of aeroderivatives with free compressors, such as Layout (a), the compressor must
absorb the power of the turbine which drives it. The compressor power per unit of air mass flow rate is
proportional to T1 and increases with the pressure ratio p2/p1. The turbine’s power per unit of mass flow
rate is proportional to T3 and increases with the pressure ratio p3/p3P. A reduction in ambient temperature
increases airflow and raises the cycle pressure ratio, just as for a single-shaft machine, causing p2 and p3
to rise in the same proportion. If the pressure ratios across the compressor and across the GGT were to
rise in the same proportion, the cooler T1 and constant T3 would reduce compressor work relative to GGT
work, and the surplus turbine work would therefore accelerate the gas generator spool. A new
equilibrium speed is established, however, at a slightly higher speed and correspondingly higher airflow,
because p3P also increases due to the greater mass flow rate being forced through the flow resistance of
the next downstream turbine spool, the PT. The flow resistance of any turbine spool is similar to that of a
choked nozzle, unless the number of stages are small (one or two) and subsonic. Thus, the pressure ratio
across the GGT, p3/p3P, does not increase as much as the pressure ratio across the compressor, p2/p1,
thereby reducing the surplus work on the free spool and it tendency to overspeed.
6-21
 M. A. Elmasri, 1990-2008
Chapter 6: GT Installation & Environment
Effect of Ambient Temperature
Typical Aero-Derivative GT
15
25
35
45
1100
110
1050
100
1000
90
950
o
5
F
-5
o
-15
580
560
540
520
500
80
900
70
850
60
800
480
Exhaust Temperature
Flow, Power, Heat Rate, % of Nominal ISO
-25
120
C
o
Compressor Inlet Temperature, C
460
440
-20
0
20
40
60
80
Compressor Inlet Temperature, oF
100
120
Flow Rate
Power
Heat Rate
Exhaust Temperature
4"/10" (10mb/25mb) Inlet/Exhaust Losses
Figure 16. Typical correction curves for an aeroderivative gas turbine (with shaft layout (a) of Fig. 15)
Due to the inverse variation of compressor speed with ambient temperature, as long as T3 is held
constant, air mass flow rate will increase as the ambient gets cooler at a rate faster than if the compressor
were compelled to rotate at a fixed speed. Thus, the correction curves for airflow, power output, and
efficiency tend to be steeper than those for single-shaft heavy duty machines.
The advantage of cooler ambients, however, is limited by the maximum tolerable speed of the compressor
spool. This may be a mechanical strength limit or an aerodynamic limit. For most aeroderivatives it is
the latter. As the ambient gets cooler, the speed of sound in air is reduced, simultaneously with the
increase in compressor speed and flow velocity, so the Mach number at compressor inlet increases. If it
approaches unity in subsonic blade passages, shock wave - boundary layer interactions could precipitate
compressor surge. Therefore the control system is programmed to reduce T3, and thereby compressor
RPM and airflow, to prevent such a limit from being reached. Once the ambient becomes cold enough to
trigger this control safeguard, the correction curves change in slope, as seen at about -7°C (20 °F) in
Figure 16. At ambients below this point, the machine is operated on a T3 below the turbine’s temperature
capability, governed by the maximum allowable corrected speed. In this choke-limited regime, the
compressor spool runs at a constant corrected speed, with a constant corrected airflow. Physical speed
and airflow therefore vary like 1/T10.5. At ambients below this point, T3 must be progressively reduced as
the ambient gets colder, resulting in a steep decline in T4, and a drop, rather than an increase, in power
output. Peak and base load power outputs are the same in the choke-limited regime.
At warmer ambients, the compressor is unchoked. The machine is operated at its T3 limit, and
compressor corrected flow and corrected speed are below their limiting values. Peak load power output
can be higher than base load power in the unchoked regime.
For aeroderivatives employing Layout (b) in Fig. 15, the correction curves are generally similar to those
shown in Fig. 16, except that they may have two break-points in their slopes, since there are two free
6-22
 M. A. Elmasri, 1990-2008
Chapter 6: GT Installation & Environment
spools that may overspeed. For some, the maximum power is at just about the 15 °C (59 °F) ISO ambient
temperature, falling on both sides.
Aeroderivatives employing Layout (c) in Fig. 15, have a flatter airflow curve, since the LPC, which sets
the airflow, runs at constant speed. Due to idiosyncrasies of certain designs, however, their correction
curves can be quite complex due to the machine encountering other limits, specific to its construction
materials and dimensions. For instance, due to the high pressure ratio, at hot ambients T2 may reach a
limiting value for the HPC materials and casing; whereas at cold ambients p2 may reach a limiting value.
These conditions can only be relieved by reducing firing temperature or by reducing airflow to reduce
pressure ratio. Reducing airflow may be accomplished by bleeding-off some compressed air from
between the LPC and HPC, or by IGV closure if the machine is so equipped.
When comparing heavy duty and aeroderivative options, the results illustrated in Fig. 16 must be borne in
mind if one is using annual average ambient for a quick estimate of generating capacity. The heavy duty
machine will perform better in winter and worse in summer so that the total kWh produced per year may
not be too far from that corresponding to the mean annual temperature. The aeroderivative, on the other
hand, may perform worse in winter and worse in summer, so an estimate based on the mean annual
condition may be misleading, particularly if that mean condition is close to its peak performance.
When using ambient temperature correction curves, it is important to differentiate between a curve at
constant specific humidity (absolute water vapour content, about 1% by volume at ISO conditions) and
one at constant relative humidity (water vapour content as a percentage of saturation content at the current
temperature). The latter can imply a very high specific humidity at hot ambients.
6.4
EFFECT OF AMBIENT HUMIDITY
A gas turbine operated with a fixed firing temperature will produce more power on a humid day than on
a dry day. This may appear counter-intuitive, since humid air is less dense than dry air, resulting in a
reduction in mass flow rate through the engine on a humid day. However, the mole flow rate (or volume
flow rate) pulled in by the compressor is independent of humidity, so on a humid day the engine has the
same mole throughflow as on a dry day. Thus, had the properties per mole of water vapor been the same
as for nitrogen and oxygen, the engine would have produced the same power, regardless of humidity. But
since water vapor has a larger heat capacity per mole than nitrogen and oxygen, with that difference
increasing with temperature, compressor work, combustor heat addition, and turbine work, all will
increase per unit airflow. Furthermore, the latter will increase more than the former. Thus, engine power
actually increases with humidity, in spite of the decreased air mass flow rate, assuming a fixed T3.
Because fuel heat input also increases, at a rate that exceeds the increase in power output, heat rate
increases and efficiency declines. Figure 17 shows these trends, at a 30 °C (86 °F) ambient.
6-23
 M. A. Elmasri, 1990-2008
Chapter 6: GT Installation & Environment
Effects of inlet air Relative Humidity on a GT Cycle with a nominal pressure ratio
of 14 at a constant firing temperature T3 of 1150 C. Ambient T1=30 C
2.00
+8 C
1.50
+6 C
1.00
+4 C
0.50
+2 C
0.00
0
Change in Parameter, %
Texhaust
Power Output
-0.50
Heat Rate
Exhaust Mass Flow Rate
Exhaust Temperature
-1.00
Fuel Input
-1.50
0
20
40
60
80
100
Inlet Air Relative Humidity, %
Figure 17. Typical effects of ambient humidity at 30 °C (86 °F) on a GT cycle with constant T3.
In spite of the minor increase in mole flow rate entering the turbine nozzles, due to the additional fuel
consumed with humid air, cycle pressure ratio decreases slightly due to the change in gas properties at the
turbine inlet. This, on its own, would lower T2 and raise T3. Additionally, the increased H2O content
lowers the specific heat ratio (γ=Cp/Cv), decreasing the exponents xC and xT of equations (6a) and (6b), so
that higher humidity results in a decreased temperature ratio across both compressor and turbine, for a
given pressure ratio, further reducing T2 and raising T3 as humidity increases.
The humidity effect depends strongly on ambient temperature, since the difference in water vapour
content between dry air and 100% humid air increases sharply with temperature. The calculations in
Figure 17 are at 30 °C (86 °F), at which 100% humidity corresponds to a water vapour content of 4.2%
(by volume or moles). On a cool 5°C (41 °F) day, 100% Relative Humidity corresponds to just 0.86%
H2O in the air, but on a hot 45 °C (113 °F) day, 100% Relative Humidity corresponds to 9.5% H2O in the
air (by volume or moles).
It should be re-emphasized that the trends discussed above, and the resulting increase in net power with
humidity, are based on assuming all other boundary conditions are the same and the firing temperature,
T3, is unchanged. In reality, firing temperature of real gas turbines may not remain constant with varying
humidity. The methods of firing temperature control are discussed later, in Chapter 7. All of them
actually control the exhaust temperature to a target value that is adjusted as a function of other measured
quantities, which may not include an adjustment for measured inlet humidity Since increasing humidity
increases the exhaust temperature, T4, for a fixed firing temperature, T3, a control system attempting to
maintain T4 at the same target value on a humid day as on a dry day will inadvertently reduce T3. This
unintentional reduction in T3 reduces power output and negates the tendency of humidity to increase
power output.
Figure 18 shows the trends reported by several manufacturers in their performance specs for various
engines. The curves for the GE Frame 6 and Westinghouse 501 show no effect or a minor reduction in
power output with humidity, indicating that the insidious drop in firing temperature with humidity must
have been taken into account when these curve were generated. The curves for the ABB and Siemens
examples show an increase in power output with humidity, indicating an assumption of constant firing
temperature in the calculations by which those curves were generated. Whether firing temperature is
indeed maintained constant by adjusting the target T4 for varying humidity is uncertain.
6-24
 M. A. Elmasri, 1990-2008
Chapter 6: GT Installation & Environment
Power Correction, % of Nominal ISO
Example of the Effect of Relative Humidity on Output
of Various Gas Turbines as Stated by their Manufacturers
100.4
100.3
100.2
100.1
100.0
99.9
99.8
99.7
99.6
99.5
0
20
40
60
80
Relative Humidity, % (59°F/15°C)
100
ABB GTX100
KWU V94.3A
W 501D5A
GE 6561B
Figure 18. Correction curves for ambient humidity for various gas turbines made by different manufacturers
6.5
EFFECT OF AMBIENT PRESSURE AND ALTITUDE
Figure 19 illustrates the impact of reduced ambient pressure, at constant ambient temperature, on the gas
turbine cycle’s T-s diagram. The solid lines 1-2-3-4 show the cycle at normal ambient pressure, and the
dashed lines 1’-2’-3’-4’ show it with a 30% reduction in ambient pressure. All corrected flows, pressure
ratios, and temperature ratios, are unchanged, so exhaust temperature is unaffected. Since inlet air density
is reduced, the air mass flow rate through a given gas turbine engine is also reduced in direct proportion to
density. Power and fuel mass flow rate are therefore reduced, in direct proportion to air mass flow rate,
and efficiency is essentially unchanged. In practice, a very small drop in efficiency is likely, since the
small mechanical losses in bearings are not reduced in proportion to airflow and power output.
Atmospheric pressure declines dramatically with altitude (as anyone who has hiked up a mountain can
tell!). The rate of pressure decline near sea level is about 1.2% per 100 m (about 3.5% per 1000 ft) of
increased elevation, and the rate of fall-off decreases with altitude, so that at 1000 m (3281 ft) pressure is
about 11% lower than at sea-level, and at 3000 m (9843 ft) about 31% lower than at sea-level. Thus,
installing a gas turbine at high altitude deprives it of a significant portion of its full capacity.
Fortunately, however, ambient temperature also falls with altitude. An isentropic atmosphere model,
would give δTa/Ta ≈ 0.286 δpa/pa, so that if sea-level were at the ISO temperature of 288.15 K (518.67
°R), ambient temperature would drop by roughly 1 °C per 100 m of elevation (about 5 °F per 1000 ft). In
reality, the atmosphere is not isentropic, and the fall-off rate is lower, about 0.65 °C per 100 m (3.5 °F per
1000 ft) near sea level, and if sea level were at ISO conditions.
Thus, at a given latitude, installing a gas turbine at high altitude would incur an output loss of about 1.2%
per 100 m of elevation, due to the lower ambient pressure. This is partly offset by the output gain of
about 0.5% per 100 m elevation, due to the cooler ambient temperature, which drops at about 0.65 °C per
6-25
 M. A. Elmasri, 1990-2008
Chapter 6: GT Installation & Environment
100 m elevation (according to equation (6a)). Additionally, the cooler ambient temperature improves GT
heat rate, whereas the lower ambient pressure does not degrade it.
Effect of Altitude on Cycle T-s diagram (illustration shows 30% reduction in ambient
density)
3'
3
1200
1100
1000
900
Temperature, C
800
700
600
500
2
400
4
2'
4'
300
200
100
0
1
-100
-0.1
0
1'
0.1
0.2
0.3
0.4
0.5
0.6
0.7
0.8
0.9
1
1.1
Specific Entropy, kJ/kg-K
Figure 19. Thermodynamic effects of reduced ambient pressure on T-s diagram
One method that may be considered to recover the lost output in a high-altitude installation is to install a
supercharging fan. This is discussed in Chapter 13.
6.6
6.6.1
GAS TURBINE FUELS & EFFECT ON PERFORMANCE
NATURAL GAS
Gas turbines operate best on clean, dry natural gas fuel. Most standard machines will operate
satisfactorily with a gaseous fuel volumetric heating value of between roughly 28,000 and 41,000 kJ/scm
(750 and 1100 BTU/scf). Operating with fuels outside this range is possible, but generally requires
special modifications to the standard fuel nozzles and other aspects of the fuel control and combustion
system, which must be made by the manufacturer. Gaseous fuel should have a consistent heating value,
and its variation during operation should not exceed ±10% without special considerations. DLN
combustors which pre-mix air and fuel may have additional constraints on fuel composition, to ensure
that flame propagation speed is within the range for which the system is designed.
Typical pipeline natural gas usually consists largely (75-85%) of Methane (CH4). The remaining 15-25%
are typically Ethane (C2H6), Propane (C3H8), Butane (C4H10), and Pentane (C5H12), with traces of heavier
hydrocarbons. Some Carbon Monoxide (CO), and some diluent gases, such as N2 and CO2 are usually
present as well. If the crude gas contains sulphur, usually as H2S, it is removed by chemical scrubbing
prior to piping the gas to the users.
If the gas contains heavier hydrocarbons (C6 and higher), they can condense at high pressure and ambient
temperature, so the gas must be heated to ensure that it is all entirely gaseous, prior to use. Fine liquid
hydrocarbon droplets entrained in the gas cause combustor vibration as they burn, shortening the life of
the combustor and transition piece. Since compressor discharge air is usually hotter than the auto-ignition
temperature of many liquid hydrocarbons, they are prone to burn prematurely in dry low NOx systems
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where air is pre-mixed with fuel prior to combustion. This can trigger flame flashback into the pre-mix
zone, with potentially dangerous consequences. Sizeable hydrocarbon liquid slugs entrained in the gas
can be very dangerous, since their arrival at the fuel nozzles would cause sudden spikes in combustor
temperature, which may cause engine failure or explosion. It is therefore necessary to ensure that all
hydrocarbons, including the traces of heavier compounds, are in the superheated vapour phase, with a
minimum superheat of about 30 °C (54 °F). Since heating the fuel using low-grade recovered heat in a
combined cycle also improves efficiency, as described in Chapter 13, it is strongly recommended.
Natural gas should be dry, since water in the gas can combine with hydrocarbons to form solid, crystalline
hydrates, especially at high pressures and low temperatures in the pipeline. To prevent this, water partial
pressure must be kept very low, which limits the permissible H2O content to less than about 150 ppm by
weight. The gas producers normally remove the H2O prior to sending the gas through the pipeline, to
prevent hydrate formation which would plague the pipeline compressor stations and valves. This may be
done by chemical scrubbing with methanol or ethylene glycol.
Natural gas can contain solid contaminants, such as soil or salts from the gas well and such as rust, scale,
or welding and grinding debris, from the pipeline. These particles can cause erosion of the fuel nozzles
and especially of the control valves, resulting in fuel flow maldistribution between nozzles. All solid
particles larger than 10 microns must be removed through a fine filter before supply to the gas turbine,
and piping downstream of the filter should be made of stainless steel and properly cleaned prior to startup
of the plant. Additional, temporary filters are recommended during the commissioning phase of a plant
with a new pipeline spur and new fuel piping.
It is usual to install a gas cyclone scrubber or inertial separator upstream of the gas filter and the fuel
compressor (if used). These devices consist of a vertical cylindrical pressure vessel, with the gas entering
at the top and travelling downwards at high speed through cyclones or nozzles. The acceleration pressure
drop cools the gas and helps to condense water vapour and liquid hydrocarbons that are near their
saturation temperature. The gas is then forced to change direction, and travel upwards behind a baffle to
the exit pipe from the tank. The inertia of the liquid droplets and solid particles causes them to continue
to travel downwards, where they collect in the pool of separated liquid at the bottom of the tank, and are
periodically blown down. Coalescing filters may be installed as a second stage of liquid removal and gas
cleanup.
Gas turbines usually need a gaseous fuel pressure well above that of their combustor, to allow accurate
fuel flow rate control by the valves. Older gas turbines without dry low NOx combustors typically
require a pressure 30-40% above the GT compressor nominal discharge pressure, so that a gas turbine
with a nominal pressure ratio of 12, say, would need a fuel pressure of about 16 bars (≈ 235 psia). Newer
gas turbines with dry low NOx combustors tend to require higher fuel pressure for sound control and
distribution, up to 65% above nominal compressor discharge pressure. Thus, a modern “F-class” gas
turbine with a pressure ratio of 16 would need a fuel pressure of about 27 bars (≈ 390 psia).
If the pipeline pressure is higher than needed by the gas turbine, a reducing station is needed to drop the
pressure. Due to the Joule-Thompson effect, throttling also drops the temperature at constant enthalpy, so
if the pipeline is at atmospheric temperature, the gas will be cooler than ambient after the reducing
station, which may cause condensation of water and/or the heavier hydrocarbons. To prevent
condensation, the gas may be heated before the reducing station, either with electric heaters or with fired
heaters. Electric heaters are simpler and have a lower capital cost, but their use has a negative impact on
overall plant net efficiency. In a large plant with appreciable pressure reduction, it may be worthwhile to
employ turbines in the reducing station, rather than just throttle the gas and waste its pressure.
Heating the gaseous fuel increases its energy content and so reduces GT fuel consumption. If the heat
were free, it would improve plant efficiency. However, the heat is not free, and the impact on plant
efficiency depends on the source of heat. If low grade heat is taken from the boiler, plant net heat rate
improves. If fuel is burnt in a fired heater, plant net heat rate becomes worse, by a negligibly small
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increment, since, except for the losses in the fired heater, the enthalpy of the fuel it consumes is
transmitted to the gas turbine. If electricity is used to heat fuel, plant net heat rate becomes much worse,
since each unit of electricity generated costs about two units of combined cycle fuel. Table 2 shows the
approximate impact of gaseous fuel heating from different energy sources on the net heat rate of a typical
dual-pressure combined cycle, fed by a high-pressure gas pipeline at 25 °C (77 °F).
Table 2. Effect of fuel heating energy source on a combined cycle’s heat rate
Source of fuel heat
Approximate Effect on Net CC Heat Rate, %,
for a 100 °C (180 °F) fuel temperature rise
Economiser water from HRSG
-0.50%
Fuel-fired heater
+0.05%
Electric Heater
+0.50%
If the pipeline pressure is too low for the gas turbine, a fuel compressor is needed. The heat of
compression will normally be sufficient to elevate the gas temperature to well above the hydrocarbon
dewpoint, so that further fuel heating may be employed solely as an efficiency enhancement in a
combined cycle, as explained in Chapter 13.
Fuel compressors are large, expensive items of equipment. In many cases, two 100% capacity
compressors are specified, to ensure plant availability if one requires maintenance. In such cases, and in
conjunction with high-pressure-ratio aeroderivatives which need a high gas pressure, the compressors’
capital cost may be as much as 10% of the gas turbine’s. They may also consume as much as 2% of the
gas turbine’s power output.
Fuel compressors are of two basic types: (a) centrifugal and (b) positive-displacement. Centrifugal
compressors are used when the flow rate is large compared to the pressure rise. They are commonly used
for large flow capacities, such as for pipelines and for boosting pressure at large power plants. Positive
displacement compressors are used when the flow rate is small compared to the pressure rise. They are
commonly used as boosters in small or medium power plants (gas turbines below 150 MW, say).
Positive displacement compressors may be of the reciprocating type or of the screw type. Reciprocating
compressors are much like internal combustion engines (in the opposite sense), with a crankshaft driving
a piston in a cylinder, and a camshaft operating the valves. The suction valve opens as the piston travels
downwards then closes at bottom dead center. The delivery valve opens just before top dead center, after
the piston has compressed the gas, then closes as the piston reverses. Screw compressors trap the gas at
the suction port between two helical screws and force it towards the delivery port in a continuous motion.
Figure 20 illustrates their principle, and Figure 21 shows a typical screw compressor skid.
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Chapter 6: GT Installation & Environment
Figure 20. Screw Compressor (Courtesy of Kobelco Co.)
Figure 21. Screw Compressor for a 25-MW-class Pratt & Whitney FT8 Gas Turbine (Courtesy of Kobelco Co.)
6.6.2
OTHER GASEOUS FUELS
6.6.2.1 Syngas from Coal Gasification
Synthetic gas from coal gasification consists primarily of CO and H2, and secondarily of CH4, as the
active fuel compounds; diluted with CO2, H2O and N2, which have no heating value. The heating value
per unit volume of both CO and H2 is about one-third that of Methane, the primary constituent of natural
gas. Due to this fact, and to the large proportion of inert diluents, syngas volumetric heating value is
much lower than typical natural gas. Syngas heating value is in the range 4,000-12,000 kJ/scm (110-320
BTU/scf), with the lower end of the range associated with air-blown gasification processes and the upper
end associated with oxygen-blown gasification processes. Syngas fuel from air-blown processes is highly
diluted with N2, at about 45% by volume, in addition to the CO2 and H2O diluents. Syngas from oxygenblown processes contains only a small percentage of N2, but the CO2 and H2O diluents are present in a
greater proportion, at about 30% by volume.
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Because of its low volumetric heating value, a large volume of gaseous fuel must be supplied to the gas
turbine combustor to accomplish a given energy input and temperature rise. This requires non-standard
combustion hardware, with larger fuel nozzles and valves. Additionally, the greater volume of
combustion gases entering the turbine require a larger first stage nozzle flow area, otherwise the pressure
ratio of the gas turbine engine would rise and pose the risk of compressor surge. This problem is
minimised if the air used for gasification were bled from the GT compressor’s discharge, so that much of
the syngas volume is taken from the gas turbine’s own airflow, rather than being added from the outside.
Nevertheless, the manufacturers have to modify the gas path to some extent, as well as the combustor
hardware, to adapt standard machines to efficiently burn syngas. The alternative to modifying the gas
path would be to lower turbine inlet temperature, but this reduces efficiency.
Syngas from air-blown processes may have a heating value per unit volume as low as 10% of natural gas.
This means the fuel flow into the combustor needs to be about ten times that of natural gas. With its
higher heating value, 25-35% that of natural gas, syngas from oxygen-blown gasifiers requires less
extensive modifications to a proven combustion system.
Besides the combustion and gas path modifications needed for syngas, the additional power resulting
from the higher volumetric flow of combustion products may also require a larger generator, as well as
mechanical modifications to handle the greater power. Since such major modifications are too expensive
unless many units were to be produced, the firing temperature is de-rated in these applications to keep
within the mechanical limits of the machine, and to minimise the extent of combustor and turbine gas path
re-design. Derating the firing temperature reduces performance below its theoretical potential with low
heating value syngas.
6.6.2.2 LPG and Gases with Heavier Hydrocarbons
LPG (Liquefied Petroleum Gas) usually consists primarily of propane and butane, with some heavier
hydrocarbons. It must be well superheated after evaporation to ensure that all its components are in the
gaseous phase. Heavier hydrocarbons have a lower heating value per unit mass than methane (CH4), but
are also much denser, so their heating value per unit volume is much higher. For example, propane
(C3H8) has a heating value per unit mass that is 8% below methane’s, but has a heating value per unit
volume that is 2½ times methane’s. Butane (C4H10) has a heating value per unit mass that is 9% below
methane’s, but has a heating value per unit volume that is more than triple methane’s. Thus, gases with a
larger proportion of the heavier hydrocarbons than natural gas will result in smaller volumetric flow rates
to accomplish a given energy input and temperature rise in the combustor. This either requires special
modifications to the fuel system, or, alternatively, the fuel may be diluted with an inert gas, such as steam,
to lower its volumetric heating value.
6.6.2.3 Hydrogen & H2-Rich Fuels
Hydrogen is at the opposite end of the spectrum to heavy hydrocarbons. Its heating value per unit mass is
very high, almost 2½ times methane’s, but its density is so low, that its heating value per unit volume is
less than one-third of methane’s. Thus, a gas turbine burning H2, or a H2-rich gas, such as refinery byproduct gas, will require much greater fuel volumetric flow rates than one burning natural gas. This
requires larger fuel nozzles and valves.
Gas turbines with conventional combustors can operate on hydrogen, subject to the necessary
enlargement of the fuel forwarding system, but gas turbines with pre-mix DLN burners generally cannot.
This is because the flame propagation speed in H2 is very fast, and would require extensive re-design and
re-development of the pre-mix DLN burners. This would not be economically feasible unless a large
number of machines employing this system were envisaged, perhaps in a futuristic “hydrogen economy”!
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Chapter 6: GT Installation & Environment
6.6.3
THERMODYNAMIC EFFECT OF GASEOUS FUEL HEATING VALUE
o
15
0
0
12
-2
-1
9
-4
6
-6
3
-8
-2
-3
-4
Exhaust Temperature Change
P.R., Power, Efficiency changes, %
o
F
Typical Heavy Duty GT
C
Effect of Fuel Heating Value
-5
0
-10
0
Nominal: P.R. = 14.1
Fuel: CH4/N2 mixture
20
40
60
Fuel LHV / CH4 LHV, %
TRIT = 2075 oF (1135 oC)
o
TET = 960 F (516 oC)
80
100
Pressure Ratio
Power
Effciency
Exhaust Temperature
Figure 22. Example showing the effect of fuel heating value on typical heavy duty GT performance
A machine running on a gas with a low or medium heating value will show improved performance. The
additional fuel volume flowing through the combustor and into the turbine nozzles will raise pressure
ratio & depress exhaust temperature. This, coupled with the increase in turbine flow relative to
compressor flow increases power and efficiency, as shown in Figure 22. The improved GT performance,
however, is illusory, since the voluminous gaseous fuel must be compressed elsewhere, absorbing power
elsewhere.
If the fuel's heating value were very low, the compressor pressure ratio could be pushed up too close to
the surge line. It may be reduced by de-rating firing temperature, but this is undesirable due to the loss of
efficiency. Closing the IGV's to reduce airflow is not necessarily viable, since the surge pressure ratio
will also be lower at the reduced airflow. One solution with coal gasifiers is to integrate the GT with the
gasifier that takes compressed air for gasification from the GT compressor, lowering the net addition of
volume to the gas turbine and thereby preventing an excessive increase in its pressure ratio.
6.6.4
DISTILLATE OIL
As its name implies, this is distilled from crude, i.e. it is condensed from vapours in the refinery
distillation column, below kerosene. This makes it clean, and free of ash or solid impurities and
essentially free of water or mineral salts. Sometimes known in the USA as No. 2 oil, it is the preferred
liquid fuel for a gas turbine, but is relatively expensive. Kerosene, the next lighter component in the
refinery column, is also an excellent gas turbine fuel.
Distillates flow freely and atomise well without the necessity of being heated, unless the storage tank
were very cold. They require very little treatment, usually just settling and filtration, with heating in cold
climates.
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6.6.5
RESIDUAL OIL AND CRUDE OIL
Residual oil, as its name implies, is the residue at the bottom of the crude oil distillation column in the
refinery. It contains the same impurities as crude oil, but in concentrated form, and is therefore the least
desirable liquid fuel for a gas turbine. It is inexpensive and sometimes the only fuel available for power
generation, since it is of low value in other applications. It is sometimes known in the USA as No. 6 oil
or as Bunker C oil.
Besides needing treatment to remove harmful contaminants, heavy oils introduce additional operational
issues. Due to their viscosity, they are hard to atomise in the fuel nozzles, and do not flow freely in the
fuel piping and flow distribution system, tending to gel up when cold, especially if they have a significant
wax content. Therefore they must be heated, to 50-125 °C (120-260 °F), to be handled properly by the
fuel forwarding, control and distribution system and for proper atomisation. The machine may be started
up and stopped with a light oil to prevent the heavy oil from gelling within its fuel system when it cools
down.
The most harmful contaminants normally present in typical crude oil are trace amounts of Sodium,
Potassium, Vanadium, and Lead, which form molten ash that adheres to turbine buckets and corrodes
them at high temperature. The sticky molten ash can also plug film cooling holes, and the solidified
deposits can alter the aerodynamic shape of the buckets. Typical fuel specs allow less than 1 ppm of each
of these metals in the gas turbine’s fuel. Ash from Calcium is not corrosive, but forms deposits, and its
allowable content is higher, generally limited to 10 ppm. Although these limits are so small that they
appear meaningless, it should be noted that a 180-MW class gas turbine consumes about 12 kg/s of fuel,
or about one million kg per 24-hour day, so with 1ppm of a contaminant in fuel, about one kg passes
through the machine per day.
Most of the Sodium and Potassium salts found in crude oil are water soluble, such as sodium chloride
(NaCl), common in oil extracted from the ocean floor or from geological formations which used to be
oceans in pre-historic times. Water-soluble salts can be removed by a washing process, in which the oil is
heated and thoroughly mixed with hot water, at about 5-10% of the oil, then passed through a centrifuge.
Since water is heavier than most oils, the centrifuge expels the water (and the salts dissolved in it), as well
as any solid particles heavier than oil, to its outer diameter, where it can be separated.
With highly contaminated oil, the water-washing process may be repeated in stages. With lightly
contaminated oil, the natural water content may be sufficient to entrain all the dissolved salts, so that
heating and centrifuging may be done without adding new water.
The water removed from the centrifuge contains traces of oil, and its disposal presents environmental
difficulties. In many cases, this too needs treatment, such as bacterial digestion of the oil traces, before
discharge to the environment.
The centrifuge can separate water as long as oil is sufficiently lighter, with a specific gravity of less than
0.98, say. If the oil is heavier, it may need blending with a lighter oil, such as a distillate, to lower
specific gravity. Alternatively, an electrostatic process may be used instead of a centrifuge.
Vanadium in the form of organic compounds cannot be removed by washing, and if present, its harmful
effects are inhibited by doping with a Magnesium salt. This reacts with the Vanadium Oxide formed
during combustion to form Magnesium Vanadate, which has a higher melting temperature, and thus does
not stick to the buckets and corrode or plug them.
Traces of Lead cannot be economically removed, so the fuel should have less than 1 ppm to begin with.
Sulphur is also corrosive, but more so in the heat recovery boiler, at lower temperatures, where vapours of
sulphur trioxide condense and form sulphuric acid. Most gas turbines can burn oil that contains up to
0.5% sulphur without difficulties. Distillate usually contains less than 0.2%, and although crude oil may
contain over 3%, most of the sulphur is removed during the washing process.
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Chapter 6: GT Installation & Environment
Solid particles in the oil should be eliminated by filtration prior to use, to prevent abrasion of the fuel
pumps, flow dividers, fuel valves and fuel nozzles.
6.6.6
PERFORMANCE DIFFERENCES BETWEEN OIL AND NATURAL GAS FUELS
Because natural gas has a higher hydrogen/carbon ratio than oil, its combustion products have a higher
percentage of H2O and a lower percentage of CO2 than the combustion products from oil. For instance,
the exhaust from a modern gas turbine at ISO conditions will have about 9% water vapour by volume
when burning gas, but only about 6% when burning oil. The specific heat of H2O is much higher than
that of CO2, so the specific enthalpy of the combustion products from burning gas is higher than from
burning oil, if both are at the same temperature. Thus, for given values of T2 and T3, a machine burning
oil will have a lower enthalpy h3, than one burning gas, and will therefore consume less oil on a LHV heat
input basis than it would consume gas. It will also produce less power than when burning gas, since the
enthalpy of the stream entering the turbine is less. The decline in power exceeds the reduction in heat
input, so the machine’s efficiency falls, but not by as much as its power output.
Because the heating value of oil is lower per unit mass, more oil is consumed by mass, which results in a
larger mass flow rate of combustion products. The combustion products with oil, however, are denser, so
their volumetric flow rate (or mole flow rate) is lower, leading to a reduced cycle pressure ratio and a
hotter exhaust temperature. Despite the larger mass flow rate and hotter exhaust, and because the exhaust
gases have a lower specific heat, they will generate slightly less steam in a heat recovery boiler, and
therefore less power in the steam turbine of a combined cycle.
Table 3 summarises typical performance, when oil is burnt, relative to performance when gas is burnt.
The calculations are done for a typical, medium heavy duty gas turbine with a nominal firing temperature
of 1180 °C (2160 °F), and a dual-pressure bottoming cycle. Natural gas is assumed as methane, and
distillate is assumed with an H/C ratio of 2 and a LHV of 18,200 BTU/lb.
Table 3. Changes in performance parameters when typical distillate oil is used as a fuel, relative to performance
when methane is used
Typical value with distillate oil
relative to natural gas*
Performance Parameter
6.6.7
GT power output
-2.5%
GT fuel input, by mass
+17%
GT fuel input, by LHV energy content
-1.3%
GT heat rate
+1.2%
GT exhaust mass flow rate
+0.3%
GT exhaust mole flow rate
-1%
GT exhaust temperature
+1 °C
ST output, typical dual-pressure combined cycle
-0.5%
CC gross output, typical dual-pressure combined cycle
-1.9%
CC gross heat rate, typical dual-pressure combined cycle
+0.6%
AN IMPORTANT REMINDER ON FUEL HEATING VALUE DEFINITIONS
Two definitions of fuel heating value are commonly used. The Lower Heating Value or LHV is the heat
that would be obtained by burning a the fuel without condensing the water vapour in the combustion
products. The Higher Heating Value or HHV is the heat obtained by burning the fuel, augmented by the
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latent heat released if the water vapour created in the combustion process were all condensed. For pure
methane, HHV is 11% greater than LHV. For oil fuels, HHV exceeds LHV by 5½-7½%, depending on
composition.
It is common in the gas industry to sell fuel based on its HHV, i.e. when natural gas is quoted in the USA
at $9 per million BTU, without further qualification, it is commonly understood to be for a million BTU
based on HHV, which is only about 900,000 BTU based on LHV. Thus, this price corresponds to $10 per
million BTU on an LHV basis. Gas price in the USA is sometimes quoted in Therms, where a Therm is
100,000 BTU, on an HHV basis. Thus, 90¢ per therm implies $9/million BTU (HHV).
It is common in the gas turbine industry to quote efficiency and heat rate based on LHV, recognising the
fact that the water vapour in the combustion products will not be condensed in practice, so its latent heat
is destined to be lost. A gas turbine that is 35% efficient on an LHV basis (Heat Rate 9750 BTU/kWh) is
only about 35/1.11 = 31.5% efficient (Heat Rate 10,820 BTU/kWh) on an HHV basis.
It is good practice to always state whether efficiency or heat rate are based on LHV, or on HHV,
whenever possible. Following gas turbine industry practice, this writer prefers to use LHV, so unless
explicitly stated otherwise, any efficiency or heat rate quoted should be understood to be based on LHV.
6.6.8
FURTHER NUANCES ON FUEL HEATING VALUE DEFINITIONS
The heating value of a fuel, LHV or HHV, depends on the initial conditions of the reactants and the final
condition of the products of combustion. For example, assume methane and oxygen, both initially at
room pressure and temperature, are reacted at constant volume in a bomb calorimeter, the reaction is
CH4 + 2 O2 = CO2 + 2 H2O ........... (20)
If the products are cooled back to the initial temperature of the reactants, and the heat removed is
measured, it defines the HHV since all the H2O vapour created will be condensed (assuming, for now,
“room temperature” is low enough to condense practically all the water vapour). The measured heat
removed in this experiment, conducted at constant volume, is the constant-volume HHV.
In the above experiment, the final products’ pressure will be at about one-third of an atmosphere, since
we began with three moles of gaseous reactants (2 oxygen moles plus 1 methane mole) and ended up with
three moles of products, two moles of water, in the liquid phase occupying a negligible volume, and one
mole of carbon dioxide in the gaseous phase, occupying the entire volume. If the products were now
isothermally compressed by the atmosphere, back to their initial pressure, the work done on the products
will be negligible for compressing the liquid water, and for the single mole of gaseous CO2 the work of
isothermal compression is
Wiso,comp = RTa ln (pa/pp) .......... (21)
where
R is the universal gas constant = 8314 J/kg-mole
Ta is the atmospheric temperature (assumed 273.15 K in our example)
pa is the atmospheric pressure (=1 atm in our example)
pp is the pressure of the products after constant-volume combustion (1/3 atm in our example)
This work done on the products will result in an equal heat removal to keep them at the same temperature.
With the numerical values shown, this comes to 8.314*273.15*ln(3) = 2495 kJ/kg-mole of CO2 (or per
kg-mole of CH4 in this case). This is equal to 2495/16.04=155 kJ/kg of CH4, roughly 0.3% of its heating
value. The difference between constant-volume and constant-pressure heating values thus depends on the
fuel composition and phase, since this affects the moles of gaseous products produced per mole of
gaseous reactants. As seen above, this difference is typically small, a fraction of one percent.
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The standard datum for defining heating value is at one atmosphere and 25 °C (77 °F). The latent heat of
water at 25 °C is 2442.52 kJ/kg. Since each kg of hydrogen in the fuel produces about nine kg of water
(actually 18.016/2.018=8.92765), HHV exceeds LHV by 21,806 kJ per kg of hydrogen in the fuel. Thus,
the difference between constant-pressure HHV and LHV, per unit mass of fuel, is
HHV – LHV = (Fuel hydrogen content by weight, kg H/kg fuel) * 21,806 kJ/kg ......... (22)
HHV – LHV = (Fuel hydrogen content by weight, lb H/lb fuel) * 9,375 BTU/lb ......... (22a)
Plant heat rate, by convention, is based upon the fuel’s heating value (or chemical energy content, defined
at 25 °C), not its total energy content. If the fuel is hotter than 25 °C (77 °F), its energy content exceeds
its heating value by the sensible heat. For example, if CH4 is heated by 100 °C (from 25° to 125°), its
enthalpy increases by about 240 kJ/kg, so its additional sensible energy content is about 0.5% of its LHV.
Thus, a gas turbine consuming the same amount of energy, would consume 0.5% less fuel by mass.
Although the real heat rate of that gas turbine is unchanged, since it consumes the same amount of energy,
its apparent heat rate improves by half a percent because it uses half a percent less fuel.
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Chapter 7: GT Operating Parameters
EFFECTS OF GAS TURBINE OPERATING PARAMETERS
Content Revised October, 2002
© Maher Elmasri 1990-2002
7.1
7.1.1
STEAM & WATER INJECTION
INJECTION FOR EMISSIONS CONTROL AND FOR POWER AUGMENTATION
Steam may be injected into a gas turbine either for NOx control, or for power augmentation, or both.
Water injection is used almost exclusively for NOx reduction, and although it also augments power
output, it is seldom used for performance purposes because of its highly deleterious effect on heat rate.
When steam or water injection is used primarily for NOx control, the diluent is injected into the flame
region, to lower the oxygen concentration in the flame, thereby lowering the peak flame temperature and
the thermal NOx generation. If water is used, it also quenches the flame due to the latent heat of
evaporation it absorbs, so a given NOx reduction can be accomplished by less water than steam. Injection
into the flame region is usually accomplished by mixing the water or steam with the fuel.
Because steam injection reduces oxygen concentration in the flame zone, it increases the emissions of CO
and UHC (unburnt hydrocarbons). If the steam concentration in the flame zone is too high, it also leads
to less stable combustion.
For NOx control, the injected steam or water is most effective when mixed with the fuel to maximise its
content in the flame zone relative to the injected amount. If mixed with the air upstream of the
combustor, the resulting steam content in the flame zone would be lower relative to the total steam
injection, since some of the injected steam will find its way to the dilution zone of the combustor. Before
the advent of reliable DLN combustors, steam injection was the most common method of NOx control
with natural gas fuel, and a ratio of steam to natural gas of approximately 2:1 by mass was generally
adequate to reduce NOx to below 25 ppm. Water injection was the most common method of NOx control
with liquid fuels, with the water mixed with the oil upstream of the fuel nozzles. A ratio of water to oil
fuel of approximately 1.25:1 by mass was generally adequate to reduce NOx to below 42 ppm with
typical oil fuels.
For power augmentation, steam or water may be injected at various locations in the engine’s combustion
section, between the compressor discharge and the turbine inlet. If the injected amount is large (more
than 4% of the airflow, say) it is injected at more than one location in the flow path to avoid excessive
concentrations of the diluent in any one location. Excessive concentrations of diluent in the flame zone
can make the flame unstable and result in high concentrations of CO and UHC. Excessive injection in the
dilution zone can cause thermal stresses due to the large difference in temperature between the hot gas
and diluent. Hence, for heavy steam injection, some designs call for three injection stations, with some
diluent injected into the compressor discharge, some directly into the flame zone, and some in the dilution
zone. The net thermodynamic effect on the cycle is identical for any injection between compressor
discharge and turbine inlet, regardless of its precise location.
Steam or water injected into the gas turbine have to be pure, originating from demineralised, boilerquality treated water. Since the injected diluent is lost in the exhaust, the cost of the water consumed can
be considerable. A 100-MW-nominal gas turbine using steam injection for NOx control will consume
water at the rate of about 40 t/hr (175 gpm), or about 1000 t/day of continuous operation.
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Chapter 7: GT Operating Parameters
7.1.2
EFFECT OF STEAM & WATER INJECTION ON GT CYCLE THERMODYNAMICS
The effects of steam or water injection on gas turbine cycle thermodynamics depend on engine
parameters, such as pressure ratio, firing temperature, turbomachine efficiencies, and turbine cooling
details. They also depend on whether firing temperature is held constant or changed when steam or water
is injected. Thus, these effects will be different for different engines. The discussions given in this
section are meant to be representative of typical heavy duty gas turbines, with an emphasis on
understanding the fundamentals, but are by no means numerically precise for all gas turbines. Indeed,
calculations show that the effects of steam and water injection can vary substantially from one machine to
another.
7.1.2.1 Effects of Steam Injection
The injected steam expands through the turbine, producing additional turbine work, without requiring a
proportionate share of compressor work. Thus, if the injected steam were 1% of the airflow by mass, it
would correspond to roughly 28.5/18 or 1.6% by volume (where 28.5 and 18 are the approximate
molecular weights of flue gases and water, respectively). Thus, to first order, turbine work would rise by
1.6%, whereas compressor work would not rise at all, so assuming that the compressor consumes half the
turbine’s work leads to the approximate result that net cycle work should increase by 3.2% for each 1% of
steam injection by mass relative to engine airflow. In reality, if T3 were held constant, the cycle pressure
ratio would increase with steam injection, causing the compressor work to increase, but also causing a
further, corresponding increase in turbine work. This latter effect adds to the net power output, beyond
the 3.2% estimated above, causing it to increase by ≈ 3.5% per 1% of steam injection relative to airflow,
assuming a fixed value for T3.
Effect of Steam Injection
20
C
o
o
F
Typical Heavy Duty GT
16
15
12
6
10
8
5
4
0
0
0
1
Nominal: P.R. = 14.1
2
3
4
Steam/Inlet Air, % by Mass
TRIT = 2075 oF (1135 oC)
TET = 960 oF (516 oC)
5
4
2
Exhaust Temperature Change
P.R., Power, Heat Rate, %
8
0
6
Pressure Ratio
Power
Efficiency
Exhaust Temperature
Figure 1. Example of the effect of steam injection on gas turbine performance parameters
7-2
 M. A. Elmasri, 1990-2002
Chapter 7: GT Operating Parameters
Effect of Steam Injection
1.95
16
1.90
14
1.85
12
1.80
10
1.75
8
1.70
Exhaust H2O, mole %
Fuel/Inlet Air, % by Mass
Typical Heavy Duty GT
6
0
1
Nominal: P.R. = 14.1
2
3
4
Steam/Inlet Air, % by Mass
o
o
TRIT = 2075 F (1135 C)
o
o
TET = 960 F (516 C)
5
6
Fuel/Inlet Air
Exhaust H2O
Figure 2. Example of the effect of steam injection on gas turbine fuel input and exhaust water vapour content
Steam injection increases fuel consumption, because the steam is injected at a temperature much lower
than T3, requiring additional fuel to be burnt to raise its temperature to T3. If we assume that the steam
comes in at close to compressor discharge temperature, and note that the specific heat of steam is about
twice that of combustion products, we can conclude that each 1% of steam injection by mass, relative to
airflow, increases fuel consumption by about 2%.
It is customary in the industry to quote heat rate & efficiency for steam injected gas turbines by charging
the engine for fuel chemical energy only, ignoring the energy input of the injected steam in the
denominator of the efficiency definition. On this basis, and from the above discussion, each 1% of steam
injection by mass, relative to airflow, improves the gas turbine’s heat rate by about 1.5%, since it
increases power output by about 3.5% and increases fuel input by about 2%.
Figures 1 and 2 show the effects of steam injection on a typical gas turbine cycle, with a pressure ratio of
≈14 and a firing temperature of ≈ 1135 °C (2075 °F).
If one were to define the efficiency of a steam-injected gas turbine based on total energy input (fuel
chemical energy plus enthalpy of injected steam) it would decline with steam injection. To make an
approximate estimate of this decline, we first note that in the absence of steam injection, the fuel input is
on the order of 1000 kJ per kg of airflow (airflow is heated by roughly 850-900 °C and specific heat is
roughly 1.15 kJ/kg-°C). For a steam injection of 1% of the airflow by mass, we already found that the
power increases by about 3.5% and the fuel energy input increases by about 2% to heat the steam within
the engine. The enthalpy of a kg of injected steam is on the order of 3000 kJ/kg, so if the steam injection
were 1% of the airflow, its enthalpy content of 30 kJ per kg of airflow represents roughly 3% of the fuel
energy input. Thus, each 1% of steam injection by mass relative to airflow corresponds to an increased
energy input to the engine of 2% (additional fuel) plus 3% (injected steam enthalpy), for a total of 5%,
whilst the power output increase is just 3.5%. On this basis, steam injection would degrade gas turbine
heat rate by about 1.5% for each 1% of steam injection by mass, relative to airflow.
7-3
 M. A. Elmasri, 1990-2002
Chapter 7: GT Operating Parameters
Effect of Steam Injection on Cycle T-s diagram (inconsistent !)
3' 3
1200
1100
1000
900
Temperature, C
800
?
700
600
500
2'
2
400
4'
300
4
200
100
0
1'≅ 1
-100
-0.1
0
0.1
0.2
0.3
0.4
0.5
0.6
0.7
0.8
0.9
1
1.1
Specific Entropy, kJ/kg-K
Figure 3. Inaccurate, constant-properties representation of the effect of steam injection on gas turbine cycle
Figure 3 shows a simplistic interpretation of the effect of steam injection on the T-s diagram of a Brayton
Cycle. This figure, based on a semi-perfect gas model, correctly shows one of the main effects of steam
injection, viz the increase in cycle pressure ratio. If T3 is fixed, the additional mole flow rate of gases
being forced into the turbine result in a higher combustor pressure, with p3’ and p2’ rising with injection in
approximate proportion to turbine inlet mole flow rate, and exceeding their corresponding non-injection
values, p3 and p2.
The pressure ratio increase, all other things being equal, would decrease, T4, as suggested by Figure 3.
However, the steam injection and attendant increase in fuel flow result in a rise in H2O content of the
gases, illustrated by Figure 2 for CH4 fuel. Thus, the process cannot simply be displayed on a T-s chart
representing gases of constant composition, a liberty we had taken in other cases without obtaining
erroneous results. The additional H2O reduces the specific heat ratio, γ, and the relative change in the
exponent of isentropic expansion (γ-1)/γ is significant. For a given pressure ratio, the temperature ratio
across the turbine is diminished, which increases T4. To estimate the magnitude of the conflicting effects
of the changing pressure ratio and flue gas composition on exhaust temperature, one uses the adiabatic
expansion equation:
T4 =
T3
 p3


p 4 

xT
=
T3
r p ,T
xT
..................................... (1)
where rp,T is the pressure ratio across the turbine (rp,T = p3/p4). For an adiabatic expansion with a
polytropic efficiency ηT, the turbine exponent xT is given by
xT = ηT
(γ − 1)
γ
............................................. (2)
where γ is the specific heat ratio of the combustion gases. Now, if all parameters are assumed constant
with steam injection, except the pressure ratio and the adiabatic expansion exponent, one may write
7-4
 M. A. Elmasri, 1990-2002
Chapter 7: GT Operating Parameters
δT4 =
∂T4
∂T
δrp ,T + 4 δxT
∂rp ,T
∂xT
............................ (3)
which, after manipulation yields
δT4
T4
= − xT
δrp ,T
rp ,T
− {n(rp ,T )}δxT
..................... (4)
For a steam injection rate of 1% by moles, (δrp,T/rp,T) ≈ 0.01, and the exponent xT has been calculated to
change per 1% water vapour by volume by about δxT ≈ -0.0006. A typical value of xT is 0.21, so the first
term on the right of equation (4), which gives the exhaust temperature drop due to the higher pressure
ratio, is about -0.0021. For a cycle pressure ratio of 16, for example, the second term on the right, which
gives the exhaust temperature rise due to the change in composition, is +0.0016. Thus, the effect of
changing composition almost negates the effect of changing pressure ratio, and the model suggests an
exhaust temperature change of –0.0005(T4), for this example, so with a T4 of 850 K, the exhaust can be
expected to drop by a about 0.4 °C per 1% (mole) steam injection.
The above adiabatic model approximates lightly cooled, old gas turbines. The effects of turbine cooling
on exhaust temperature are significant with modern, heavily cooled turbines, in which the cooling air can
exceed 15% of engine airflow. When steam is injected, the higher pressure ratio leads to hotter cooling
air. Furthermore, the ratio the heat capacity of the cooling air to that of the hot gases is reduced with the
increase in hot gas mass flow rate and specific heat. Both of these effects raise exhaust temperature, T4.
Hence, to find the effect of steam injection a complete engine model is needed, including the various
details of turbine cooling. The results of such a model appear in Figures 1 and 2, which show that with all
effects considered, exhaust temperature rises with steam injection if the firing temperature, T3 (also called
turbine rotor inlet temperature and abbreviated as TRIT in Figs. 1 and 2), were held constant.
7.1.2.2 Effects of Water Injection
Water injection has a similar effect to steam injection on power output, but a substantial negative effect
on heat rate. Whereas injected steam is usually superheated, and in the temperature range of 250-450 °C
(480-840 °F), injected water is usually at a modest temperature, in the range of 15-150 °C (60-300 °F).
The additional fuel burnt to heat injected steam has to just supply the energy to superheat it from its
injection temperature to T3, but the additional fuel burnt with water injection has to supply the energy to
evaporate the water, as well as to heat it to T3, starting from a much lower temperature. Each kg of
injected steam requires additional fuel energy on the order of 2000 kJ, whereas each kg of injected water
requires additional fuel energy on the order of 4800 kJ.
Thus, whereas each 1% of steam injection relative to engine airflow increases fuel consumption by about
2%, each 1% of water injection increases it by about 5%. Either would increase power by about 3.5%.
Thus, the gas turbine’s apparent heat rate, based solely on fuel chemical energy input, improves by about
1.5% with steam injection, but deteriorates by about 1.5% with water injection.
7.1.3
EFFECT OF STEAM & WATER INJECTION ON GT BLADE METAL TEMPERATURES
When steam or water is injected into the gas turbine, the heat capacity flux, (ρVCP), the product of gas
density velocity and specific heat increases, since both the mass flux (ρV) and the specific heat (CP)
increase. Hence, the heat transfer coefficient between the gas and the turbine blades increases as well.
The effect on blade metal temperatures depends on specific design details, such as whether the blades are
internally cooled or film-cooled, the coolant temperature, and the blade cooling effectiveness curve.
While no generalisation can be considered accurate, this author’s estimate is that for typical first-stage
design parameters each 1% of steam or water injection by mass, relative to engine airflow, causes a first7-5
 M. A. Elmasri, 1990-2002
Chapter 7: GT Operating Parameters
stage blade temperature rise of about 3-4 °C (5-7 °F). With typical NOx control steam injection rates of
3% of airflow, the estimated increase in first stage metal temperatures with typical parameters is therefore
9-12 °C (16-22 °F), which would reduce blade life by a factor of between 1½ and 2, and is equivalent to a
firing temperature increase on the order of 17 °C (30 °F).
It is therefore desirable (and common) to de-rate firing temperature by about that amount when steam
injection for NOx control is applied.
7.1.4
EFFECT OF STEAM & WATER INJECTION ON BEHAVIOUR OF AERODERIVATIVES
In a single-shaft machine, steam injection does not affect engine rotational speed or airflow. The
additional mass flow rate and declining molecular weight of gases entering the turbine nozzle result in a
higher volume of gases being forced into the turbine. If T3 is fixed, this will force the cycle pressure ratio
to increase. Power will increase and heat rate will appear to improve, as shown in Figure 1. If high levels
of steam injection are contemplated, the firing temperature may need to be reduced, both to prevent blade
overheating and to limit the increase in pressure ratio to maintain an adequate surge margin for the
compressor.
For aeroderivatives in which the LPC is on a free shaft, such as shown in Figures 15a and 15b in Chapter
6, steam injection increases LPC rotational speed and hence airflow. This is because the additional mass
flow rate through the turbine stages on any free spool results in surplus turbine work relative to that
consumed by the driven compressor, causing the spool to accelerate, until a new equilibrium speed is
attained, as described in §6.3.3. The higher speed and the corresponding increase in airflow improves
performance, but is possible only if the compressor does not overspeed, mechanically or aerodynamically,
to its limiting physical or corrected speed.
To illustrate, we consider a typical aeroderivative using Layout (a) in Figure 15a, with a maximum, base
load firing temperature of 1240 °C (2264 °F) and a maximum corrected airflow of 67.5 kg/s (149 lb/s).
Firing temperatures above the maximum value would reduce hot section life, and corrected airflows
above the maximum value would correspond to excessive corrected speed, posing the risk of compressor
surge. Figure 4 shows the airflow variation with ambient, with and without steam injection. The closed
symbols show physical airflow, and the open symbols show the corresponding corrected airflow. Figure
5 shows the corresponding exhaust temperature, along with the estimated firing temperature. The
maximum firing temperature of 1240 °C applies without steam or water injection. With steam injection,
the maximum firing temperature is reduced by about 20 °C, to 1220 °C, to preserve hot section life.
7-6
 M. A. Elmasri, 1990-2002
Chapter 7: GT Operating Parameters
o
Compressor Inlet Temperature, C
-25
-15
-5
5
15
25
35
45
72
155
70
150
68
145
66
64
140
62
135
Airflow, kg/s
Airflow, lb/s
160
60
130
58
125
-20
0
20
40
60
80
100
o
Compressor Inlet Temperature, F
120
Dry
25 kpph (3.15 kg/s)
30 kpph (3.78 kg/s)
4"/10" (10mb/25mb) Inlet/Exhaust Losses
Closed symbols: Actual airflow
Open Symbols: Corrected airflow
Figure 4. Effect of steam injection on physical and corrected airflow in a typical aeroderivative
Compressor Inlet Temperature, oC
-5
5
15
25
35
45
1250
2200
1200
2100
1150
2000
1100
1050
1900
1100
550
1000
o
-15
Temperature, C
o
Temperature, F
-25
2300
500
900
450
800
400
700
-20
0
20
40
60
80
Compressor Inlet Temperature, oF
4"/10" (10mb/25mb) Inlet/Exhaust Losses
Closed symbols: Estimated TRIT
Open symbols: TET
100
120
Dry
25 kpph (3.15 kg/s)
30 kpph (3.78 kg/s)
Figure 5. Variation of estimated firing temperature with ambient for a typical aeroderivative, dry (without steam
injection) and with different levels of steam injection
At warm ambients T3 is at its maximum, and corrected airflow is below its maximum. At cold ambients,
the converse applies, i.e. the corrected airflow is at its maximum, and the controls have been compelled to
lower T3 to reduce the rotational speed and avoid aerodynamic overspeed. The control system monitors
7-7
 M. A. Elmasri, 1990-2002
Chapter 7: GT Operating Parameters
both T3 and corrected speed. Both measurements are indirect. T3 is measured by measuring PT inlet
temperature and calculating T3, and corrected speed is measured by measuring physical speed and
compressor inlet temperature and calculating corrected speed. The control logic at base load is to keep T3
at its maximum value (1240 °C without injection and 1220 °C with injection), unless corrected speed
were to exceed its maximum value, in which case fuel flow would be reduced to keep corrected speed at
its maximum value. This would cause T3 to be lower than its maximum.
Without steam injection, the aerodynamic overspeed limit prevails at ambients below about 5 °C (about
40 °F). With 3.15 kg/s (25,000 lb/h) of steam injection, the engine runs faster at warmer ambients, so the
aerodynamic overspeed limit prevails at ambients below about 15 °C (59 °F). With the greater rate of
steam injection of 3.78 kg/s (30,000 lb/h), the aerodynamic overspeed limit prevails at even warmer
ambients, below about 20 °C (68 °F).
Figure 6 shows the corresponding power output and efficiency curves.
maximum power is at an ambient of about 15 °C (59 °F).
With the steam injection,
If steam injection is for power augmentation, it should be realised that once the engine is governed by its
limiting corrected speed, there is very little benefit to GT power output by steam injection, since it must
be counterbalanced by a reduction in T3 to avoid overspeed. If the gas turbine is part of a combined
cycle, the reduction in T3 causes a corresponding drop in T4, as shown in Fig. 5, which reduces steam
turbine output, making steam injection counterproductive at cooler ambients.
Compressor Inlet Temperature, oC
-15
-5
5
15
25
35
45
40
32
38
30
36
28
34
26
32
24
30
22
28
20
26
18
24
Power, MW
Efficiency, %
-25
16
-20
0
20
40
60
80
o
Compressor Inlet Temperature, F
4"/10" (10mb/25mb) Inlet/Exhaust Losses
Closed symbols: Efficiency
Open symbols: Power
100
120
Dry
25 kpph (3.15 kg/s)
30 kpph (3.78 kg/s)
Figure 6. Variation of power output and efficiency with ambient for a typical aeroderivative, dry (without steam
injection) and with different levels of steam injection
7-8
 M. A. Elmasri, 1990-2002
Chapter 7: GT Operating Parameters
7.2
FULL-LOAD FIRING TEMPERATURE CONTROL
T3 is too high to be measured reliably and accurately. It is therefore normal to control fuel flow at full
load such that a lower, measurable temperature is controlled to a target value.
For single-shaft GT's, the control temperature is normally that of the exhaust, which is low enough to be
measured accurately by an array of thermocouples, usually 10-20 arranged around the circumference of
the exhaust, and in several radial locations. These measurements are also needed to ensure that the fuel
flow is distributed evenly around the perimeter of the machine to the combustor cans (or nozzles in an
annular combustor), and if they detect an excessive variation between the hottest and the average
circumferential location will trigger an alarm or a trip. Elaborate methods of voting and averaging these
measurements provide an accurate average exhaust temperature measurement.
For heavy duty gas turbines, the target exhaust temperature programmed into the controller may be
determined as either:
(i) a function of compressor discharge pressure (p2) or,
(ii) a function of compressor inlet temperature (T1);
in such a way as to correspond to constant (or nearly constant) value of T3.
For aeroderivatives, the control temperature is commonly that at power turbine inlet, rather than at
exhaust, but the same general logic is applied and the fuel flow is modulated for PT inlet temperature to
follow a target value. This target may be adjusted as a function of pressure ratio to attain a nearly
constant T3, or may simply be a fixed value, without any pressure ratio bias, resulting in a variable T3
within its acceptable range.
7.2.1
TARGET EXHAUST TEMPERATURE BASED ON PRESSURE RATIO
For an adiabatic turbine, equation (1) relates the temperature ratio across the turbine to the pressure ratio
across it. Thus, with such a model, we can derive a target value of T4 that results in T3 being maintained
at a fixed, given value, through the simple expression
T4,TARGET =
T3, FIXED
 p3
 xT

p 4 

..................................... (5)
Both p3 and p4 can be easily measured, directly or indirectly. p3 is measured by measuring p2 and
assuming that p3 is a fixed percentage below it. p4 can be directly measured, or assumed to be a fixed
percentage above atmospheric pressure. Thus, the control system can “know” T4,TARGET from the fixed T3
and the measured pressure ratio. The real T4 is measured by suitably averaging an array of many
thermocouples in the exhaust. When running at base load, the control system keeps T4 ≅ T4,TARGET via a
proportional control logic, so that if T4 exceeds T4,TARGET fuel flow is reduced and vice versa.
For a real turbine, the calculation is much more complex than equation (5), since stage efficiencies and
the introduction of cooling air must be considered. However, the fundamental fact that temperature ratio
across the turbine is a function of pressure ratio does not change, and a relationship between T4,TARGET and
pressure ratio can be numerically determined for full-load conditions. Such a relationship is shown in
Figure 7. The solid, nearly-straight curve on Fig. 7 shows calculated T4,TARGET for an assumed, fixed T3.
In older control systems, it was common to linearise the target curve as shown by the dotted line, which
results in a slight reduction of firing temperature at points away from the nominal design point, where the
line is tangent to the curve. The maximum reduction in T3 by following the straight line, rather than the
curve, is typically no more than 10° C (18° F ).
7-9
 M. A. Elmasri, 1990-2002
Chapter 7: GT Operating Parameters
Exhaust Temperature Control Curves
1125
585
1075
570
1050
555
1025
540
525
1000
510
975
495
950
Exhaust Temperature, oC
o
Exhaust Temperature, F
600
Maximum Tx
1100
480
925
6
8
10
Pressure Ratio
12
14
Linearized Exhaust Temperature Control
Fixed TIT No STIG
Fixed TIT with STIG
Figure 7. Typical control curve defining T4,TARGET as a function of pressure ratio to attain a fixed base-load value of T3
Fig. 7 shows a dotted curve for T4,TARGET at 4% steam injection, which if followed, would result in the
same T3 with that level of steam injection. This arises from the fact that T4 rises with steam injection at
constant T3 for a typical engine, as shown in Fig. 1. If, with 4% steam injection, the lower, non-injection,
solid curve were followed instead of the dotted curve, the resulting T3 will be reduced, offsetting the
tendency of steam injection to overheat the blades. Thus, it is common to follow the same target exhaust
temperature curve with and without injection, both for simplicity, and to automatically derate T3 as more
steam is injected, countering the tendency of injection to overheat the blades.
Similar logic applies to the effect of ambient humidity. To get the same T3 on a humid day as on a dry
day, the controller should follow a higher T4,TARGET , and, if it follows the same target, T3 will be derated,
negating the tendency of higher ambient humidity to raise power output.
It must be noted that the control system may set the T4,TARGET bias based on a live measurement of p2,
without a live measurement of exhaust pressure and computation of pressure ratio. In this case, the
relationship between p2, and pressure ratio would be assumed constant, but site-specific, based on the
mean back pressure at the site, which depends on the altitude and the nominal back pressure imposed by
the HRSG and exhaust system. In such cases, T3 will vary to some extent with barometric pressure and
the changing boiler back pressure, which depends on exhaust gas temperature, flow rate, composition, and
level of duct firing, if used.
7.2.2
TARGET EXHAUST TEMPERATURE BASED ON COMPRESSOR INLET TEMPERATURE
In this method the engine is simulated at full-load at its design T3 across the range of ambients. The
resulting T4 variation with ambient, as illustrated in Figs. 13 & 14 of Chapter 6, is programmed into the
controller as T4, TARGET. This method can also automatically compensate for the tendency of steam
injection to overheat the blades. Since steam injection at constant T3 would raise T4, the controller would
reduce fuel flow to restore T4 to its non-injection target value, thereby reducing T3.
7-10
 M. A. Elmasri, 1990-2002
Chapter 7: GT Operating Parameters
7.3
7.3.1
PART-LOAD PERFORMANCE & CONTROL
PART-LOADING SINGLE-SHAFT MACHINES
7.3.1.1 Turbine Inlet Temperature (TIT) Control
Older single-shaft, heavy duty machines have no variable inlet guide vanes, so once synchronised to their
fixed speed, their airflow is fixed for a given ambient. Power output can then only be varied by adjusting
firing temperature, through modulating the fuel flow, at constant airflow.
Load Reduction (to 70% ) by Reducing Firing Temperature @ Constant Airflow
1200
3
1100
3'
1000
900
Temperature, C
800
700
600
500
2
400
300
4
2'
4'
200
100
0
1'≅ 1
-100
-0.1
0
0.1
0.2
0.3
0.4
0.5
0.6
0.7
0.8
0.9
1
1.1
Specific Entropy, kJ/kg-K
Figure 8. Cycle diagram at part-load with TIT control
Figure 8 shows the effects on the cycle T-s diagram. The cycle 1-2-3-4 represents full-load, and the
modified cycle 1’-2’-3’-4’ represents 70% load, achieved by reducing T3 only. This reduces pressure
ratio, since p3 is essentially proportional to T3½ with the turbine behaving as a choked nozzle. Since
efficiency depends primarily on pressure ratio, it too is reduced. Exhaust temperature, T4, falls with the
declining T3. Figure 9 shows the change in main performance parameters when T3 is varied at constant
airflow. To reduce load from 100% to 75% requires a reduction in TIT of about 160 °C (290 °F), causing
the exhaust temperature to fall by about 90 °C (160 °F) between full-load and 75% load. This drop is too
large for efficient heat recovery in a combined cycle.
7-11
 M. A. Elmasri, 1990-2002
Chapter 7: GT Operating Parameters
Part Load Performance - TIT Control
o
120
200
115
100
50
110
0
0
105
-100
100
-200
95
-300
90
-400
85
-500
100
-50
-100
-150
TRIT & TET Changes
P.R., Heat Rate, %
o
C
F
Typical Heavy Duty GT
-200
-250
60
70
80
90
Power Output, %
100
110
120
Pressure Ratio
Heat Rate
TRIT
TET
o
o
TRIT = 2075 F (1135 C)
TET = 960 oF (516 oC)
Nominal: P.R. = 14.1
Air flow constant
Figure 9. Example of the changes in key cycle parameters as load is varied by TIT control
7.3.1.2 Inlet Guide Vane (IGV) Control
Load Reduction (to 70% ) by Reducing Airflow @ Constant Firing Temperature
3
1200
3'
1100
1000
900
Temperature, C
800
700
600
500
2
400
4
2'
300
4'
200
100
0
1'≅ 1
-100
-0.1
0
0.1
0.2
0.3
0.4
0.5
0.6
0.7
0.8
0.9
1
1.1
Specific Entropy, kJ/kg-K
Figure 10. Cycle diagram at part-load with IGV control
Gas turbines equipped with inlet guide vanes may accomplish load reduction by reducing the airflow,
with TIT held constant. Figures 10 and 11 illustrate the results. Pressure ratio falls, in approximate
7-12
 M. A. Elmasri, 1990-2002
Chapter 7: GT Operating Parameters
proportion to airflow, which reduces efficiency. With a constant T3, the lower pressure ratio raises the
exhaust temperature, T4. In the example of Figure 11, reducing load by 25% (from 100% to 75%)
requires an airflow reduction of about 23%, the rest of the load drop coming from the lower specific work
due to the lower pressure ratio. Exhaust temperature climbs by about 40 °C (70 °F).
Part Load Performance - IGV Control
120
120
110
100
C
o
o
F
Typical Heavy Duty GT
50
100
80
40
90
60
80
40
70
20
10
0
0
60
60
Nominal: P.R. = 14.1
TRIT constant
70
80
Power Output, %
90
o
o
TRIT = 2075 F (1135 C)
TET = 960 oF (516 oC)
30
20
Exhaust Temperature Changes
P.R., Airflow, Heat Rate, %
60
100
Pressure Ratio
Airflow
Heat Rate
Exhaust Temperature
Figure 11. Example of the changes in key cycle parameters as load is varied by IGV control
7.3.1.3 Combined TIT & IGV Control
TIT control results in a higher cycle pressure ratio for a given load than IGV control. It has a slight heat
rate advantage, besides the fact that operating at reduced temperatures increases life, so without heat
recovery, TIT control is advantageous.
For applications with heat recovery, IGV control is superior since it concentrates the GT exhaust energy
into a smaller massflow of higher temperature gases. This promotes efficient heat recovery by depressing
stack temperatures. It also ensures attainment of the desired steam temperatures for process or combined
cycles, which may not be possible with TIT control.
Practical limitations curtail the use of IGV control. First, the range of possible flow reduction using
IGV's is limited to about 25 %. Second, the elevated exhaust temperature can be problematic for the last,
uncooled, turbine stage(s). It is therefore common to combine the two methods in modern machines.
On many gas turbine models, load is reduced by closing the IGV's until they are fully closed, or the
maximum permissible exhaust temperature is attained, usually about 50 °C (90 °F) above nominal ISO
base load exhaust temperature. This would typically correspond to about 20-25% load reduction at ISO
ambient. For still lower loads, TIT would be reduced at constant airflow. Figure 12 shows an example
for a typical single-shaft GT controlled by an exhaust temperature control curve of the type shown in Fig.
7. A signal to reduce load would cause the IGV's to close while fuel flow is modulated to maintain T4 at
T4,TARGET. This keeps T3 essentially constant, as the lower pressure ratio resulting from reduced airflow
calls for a higher T4. Once the IGV's have reached their fully closed position, they would signal the
7-13
 M. A. Elmasri, 1990-2002
Chapter 7: GT Operating Parameters
control system that they cannot be closed any farther, so if the inputs call for still lower loads, fuel would
be reduced without attempting to maintain T4 = T4,TARGET, and the control curve be abandoned.
Exhaust Temperature
C
550
500
450
400
350
300
F
150
1100
1000
140
900
130
800
120
700
110
600
250
500
200
400
150
300
100
200
50
General Electric PG 6541B
o
100
90
80
100
70
0
25
50
75
100
Percent Nominal Power, %
125
Percent Nominl Exhaust Flow, %
o
Ta = 120 oF (49 oC)
Ta = 59 oF (15 oC)
Ta = 0 oF (-18 oC)
Symbols: Manufacturer's Data
Closed symbols: Exhaust Temperature
Solid Lines: GT PRO
Open symbols: Exhaust Flow
Figure 12. Exhaust temperature and airflow for a typical engine in which load is reduced by IGV closure followed by
TIT reduction. Percent nominal power is relative to full load at ISO conditions.
Some gas turbine models use both IGV and TIT modulation to maintain a constant exhaust temperature at
part load. This is usually done for machines where T4,TARGET is simply a function of compressor inlet air
temperature. Figure 13 shows typical part load behaviour for such an engine. The signal to reduce load
results in closure of the IGV's while the control system is modulating fuel to maintain the target exhaust
temperature, which is constant at a given compressor inlet temperature. Thus TIT is reduced as the IGV's
close. Further load reduction can only be accomplished by abandoning T4,TARGET, and reducing firing
temperature at nearly-constant airflow.
7-14
 M. A. Elmasri, 1990-2002
Chapter 7: GT Operating Parameters
Exhaust Temperature
C
550
500
450
400
350
300
F
150
1100
1000
140
900
130
800
120
700
110
600
250
500
200
400
150
300
100
200
50
Seimens V84.3
o
100
90
80
70
100
0
25
50
75
100
Percent Nominal Power, %
125
Percent Nominl Exhaust Flow, %
o
Ta = 59 oF (15 oC)
Symbols: Manufacturer's Data
Closed symbols: Exhaust Temperature
Solid Lines: GT PRO
Open symbols: Exhaust Flow
Figure 13. Exhaust temperature and airflow for a typical engine in which load is reduced by IGV closure
simultaneously with TIT reduction, to maintain a constant exhaust temperature
7.3.2
LOAD REDUCTION FOR AERODERIVATIVES
A "typical" aeroderivative with a free power turbine achieves part load by simply reducing fuel input.
This automatically reduces airflow since the compressor spool(s) will slow down. The combination of
TIT/airflow reduction is dictated by the work balance on the free spool. The extent of airflow reduction
may, in principal, be independently controlled to an extent by IGV/variable stator angles. This, however,
is not typical practice, since the variable vanes on the compressor are normally adjusted solely as a
function of spool RPM to optimize compressor efficiency, eliminating IGV angle this as an independent
degree of freedom. Figure 14 illustrates a typical example, showing how the exhaust temperature and
airflow fall in unison as fuel is reduced, reducing T3 and free spool RPM together.
Aeroderivatives such as the LM6000, where the LP compressor is tied to the load shaft, will have a
constant LP compressor speed, irrespective of firing temperature. Thus, their part-load behaviour
resembles single-shaft machines with pure TIT control unless independently adjustable variable inlet
guide vanes were used on the LP compressor inlet.
7-15
 M. A. Elmasri, 1990-2002
Chapter 7: GT Operating Parameters
Exhaust Temperature
C
y
o
F
1000
150
900
140
800
130
700
120
600
110
250
500
100
200
400
90
150
300
80
100
200
70
500
450
400
350
300
50
100
60
0
25
50
75
100
Percent Nominal Power, %
125
Percent Nominl Exhaust Flow, %
o
Ta = 100 oF (38 oC)
Ta = 59 oF (15 oC)
Ta = 0 oF (-18 oC)
Symbols: Manufacturer's Data
Closed symbols: Exhaust Temperature
Solid Lines: GT PRO
Open symbols: Exhaust Flow
Figure 14. Exhaust temperature and airflow for a typical aeroderivative engine in which load is reduced by reducing
fuel flow, prompting a reduction in LPC speed and airflow as T3 falls.
7-16
 M. A. Elmasri, 2000-2008
Chapter 8: Heat RecoveryBoilers
HEAT RECOVERY BOILERS
Content Revised April, 2002, Updated September, 2008
© Maher Elmasri 2000 - 2008
The Heat Recovery Boiler (abbreviated HRB) is frequently referred to as a Heat Recovery Steam Generator
(abbreviated HRSG). Both of these abbreviations are used interchangeably, to mean the same thing, in this
seminar. The HRB consists of a duct, within which the exhaust gases flow. Tubes carrying flowing water
to be boiled into steam are fixed within and across this duct.
A HRB for a modern, 180-MW-class GT, providing steam for a 100-MW-class steam turbine, would weigh
about 1750 tonnes (the weight of a WW2 Destroyer) and provide about 1 ½ million ft2 (~135,000 m2), or 32
acres (the size of a small farm) of heat transfer surface area. It would cost about $20 million +/- 20%
depending on configuration, or about $13/ft2 of heat transfer surface (~$150/m2), or about $5/lb ($11/kg) and
about $200 per kW of steam turbine output.
1. BOILER TUBING
Exhaust from a gas turbine is at modest temperature, on the order of 900-1100 ºF (480-600 ºC). This
temperature is not significantly higher than the steam temperatures used in steam turbines, so the boiler
operates with fairly small temperature differences, and thus requires large surface areas by comparison with
fired boilers of the same steam production. To keep the surface area and cost reasonable, practical designs
employ finned tubes to increase the surface area on the gas-side, which has much lower heat transfer
coefficients than the water-side, since the gas is much less dense than the water.
Figure 1. Finned tubes with solid fins (left)
and serrated fins (right), courtesy of Escoa.
Figure 1 shows typical finned tubes. They are mass-produced by turning the tube in a lathe-like machine to
wind the sheet metal that will form the fins around it in a tight spiral. As the fins are wound, they are
electric-resistance welded to the tubes. To form serrated fins, the sheet metal strip is sliced part-way
through at close intervals as it is being wound. In some cases, the sheet metal strip is bent to form an "L
shape" as it is being wound, with the short side of the "L" resting on the tube surface to form better thermal
contact and to help stiffen the tube. One manufacturer (IST) uses brazing, instead of welding, to make
finned tubes of high grade materials, such as Inconnel.
8-1
 M. A. Elmasri, 2000-2008
Chapter 8: Heat RecoveryBoilers
In practical designs, the gas velocity within the duct is in the range of 45-65 ft/s (14-20 m/s) at the HRB
inlet. As the gas cools down to the stack temperature, this velocity may drop to the range of 20-30 ft/s (6-9
m/s). These practical velocities result from consideration of gas-side pressure drop, which is usually in the
range of 6-12 inches of water (15-30 mbars). Higher gas velocities lead to excessive pressure drops, which
adversely affect the gas turbine's performance, as well as pose the potential for erosion. When typical gas
turbine exhaust flows at these velocities across banks of tubes, the resulting gas-side heat transfer
coefficients are fairly low, in the range 8-14 BTU/hr-ft2-ºF (45-80 W/m2-ºC).
Figure 1a. Tube panels with serrated fins.
The water side heat transfer coefficients in typical, practical designs, are much higher than the gas side
coefficients. The economisers pre-heating the forced flow of liquid water would typically have coefficients
around 600 – 1200 BTU/hr-ft2-ºF (3400- 6800 W/m2-ºC), or fifty to one hundred times the gas-side
coefficients. The evaporator tubes where the water boils have very high internal heat transfer coefficients,
in the range 2500 – 10,000 BTU/hr-ft2-ºF (14- 57 kW/m2-ºC), depending primarily on pressure. These
values are over two hundred times the gas-side coefficients, so their precise value does not matter
practically, since their corresponding thermal resistance, which is the inversely proportional to heat transfer
coefficient, is effectively zero. The superheater tubes carrying the steam would typically have internal heat
transfer coefficients varying from about 40 BTU/hr-ft2-ºF (230 W/m2-ºC) for a low-pressure superheater, to
as much as 800 BTU/hr-ft2-ºF (4500 W/m2-ºC) for a high-pressure superheater. These values are about four
times to eighty times the gas-side coefficients. Table 1 summarises these ranges.
Whenever possible, a well-designed heat exchanger should have a thermal resistance (1/hA) on each side of
a comparable order of magnitude. This helps cost-effectiveness by efficient utilisation of the surfaces. For
finned tubes used in boilers, the maximum practical ratio of the external, gas-side finned area to the internal
tube wall area is about 12:1. The fins are only partially effective, because their temperature is not the same
8-2
 M. A. Elmasri, 2000-2008
Chapter 8: Heat RecoveryBoilers
as the tube wall to which they are attached, but rises towards the gas temperature. Thus, the heat transfer
from the fins per unit of their area is not as strong as from the tube. Typically, their effectiveness is about
60%, so using them to increase the gas-side area by a factor of twelve corresponds to an effective area
increase by a factor of about seven. Thus, after considering the effect of the fins, a typical economiser,
evaporator, or high-pressure superheater, is still dominated by the gas-side thermal resistance, which is 7 to
30 times as much as the water-side resistance. A heavily-finned low pressure superheater, on the other
hand, may have a larger thermal resistance on its steam side than on its gas side.
Table 1. Typical HRB ratios of internal (water-side) to external (gas-side) heat transfer coefficients
and overall conductance with heavily-finned tubes
Heat Exchanger
hw / hg
(hA)w / (hAeff)g
Economiser
50 - 100
7 - 15
Evaporator
> 200
> 30
Superheater
4 - 80
0.5 - 10
1.a. Selection of Fin and Tube Geometry:
Bare tubes are used when gas-side temperatures are very high, to the point of "burning-off" the fin tips,
which are much hotter than the tube wall to which they are attached. At normal temperatures, they are not
as economical as finned-tubes, because their gas-side surface area per unit length of tubing is far smaller
than finned tubes.
Solid fins result in a slightly larger and more expensive boiler than serrated fins, but provides greater
corrosion resistance and cleanability if liquid fuels are to be used.
Serrated fins provide greater gas side heat transfer as well as pressure drop. They allow a smaller total
surface area and thus a less costly boiler. However, they are more prone to fouling and corrosion if the flue
gases are dirty, thus are normally used with clean natural gas fuel.
Staggered Tubes
In-line Tubes
Longitudinal pitch
Transverse pitch
Flow
Flow
Figure 2. Tube Layouts
In-line tubes are easier to clean, but result in a slightly larger, more expensive boiler than staggered tubes.
8-3
 M. A. Elmasri, 2000-2008
Chapter 8: Heat RecoveryBoilers
Staggered tubes give greater heat transfer coefficients and gas-side pressure drops, helping to make the
HRSG more compact and less costly. This makes them well suited for natural-gas-fired plants, but with oilfiring, if soot or other accumulation is a concern, they should not be used since they are harder to clean with
soot-blowers.
Longitudinal pitch is the distance in the gas flow direction from the center-line of one tube row to the next.
Transverse tube pitch is the distance perpendicular to the gas flow direction, from the center of one tube to
the next. In typical practice, both are about 1.3-1.4 times the diameter of the tube envelope, at the fin-tips.
1.b. Selection of Finned Tube Dimensions:
For unfired heat recovery boilers, common fin heights are 5/8-¾" (15-19 mm) and common tube diameters
are 1¼ -2" (32-51 mm) diameter tubes. Short fins give less surface but more efficiency, i.e. a high degree of
utilisation of the extra surface. Long fins add more surface with less efficiency. If high levels of
supplementary firing are used, shorter fins or even bare tubes should be used, since the tips of long fins
would overheat and distort or corrode.
Fin spacing, measured between adjacent inside surfaces, equals pitch (center-to-center) minus fin thickness:
1
number per unit length
1
− thickness
fin spacing =
number per unit length
fin pitch =
Typically, if the flue gases are clean, up to 6 fins per inch are used (pitch = 0.167 " = 4.2 mm); with dirty
gases, 3 fins per inch are reasonable (pitch = 0.33" = 8.4 mm).
With heavy supplementary firing, it may be desirable to use wider fin spacing on the superheater tubes, to
reduce heat flux, allowing the tube to stay closer to the steam temperature, rather than letting it to heat up
towards the gas temperature.
Thick fins are more efficient, abrasion-resistant, and corrosion-resistant, but more costly to make. In heat
recovery boilers, fin thicknesses between 0.03" (0.75 mm) and 0.05" (1.25 mm) are generally used.
For many years, common US practice was to use tube outer diameters of 2" (50.8 mm) for economisers,
evaporators and superheaters. This relatively large diameter is a good choice for natural circulation in the
evaporator, and for superheaters, where the volumetric flow rate is large, due to the low density of the
vapour. It is not always a good choice for economisers, which carry the much denser liquid phase, or even
for high-pressure evaporators and superheaters, where the vapour is also dense enough that such large tubes
result in low vapour velocities. Several US designs therefore use 1½" (38 mm) or even 1¼" (32 mm) tubes
for the economisers, to secure higher water velocities and internal heat transfer coefficients. The trend in the
more recent designs is to use 1½" (38 mm) tubes for the high-pressure superheaters and evaporators. Lowpressure superheaters and especially reheaters are generally made with larger tube diameters, 2-2½” (50-63
mm).
Many European manufacturers, who tend to produce vertical rather than horizontal boilers, have historically
used tube diameters smaller than 2" (50.8 mm). In addition to increasing the internal heat transfer
coefficients with a dense water-side flow, the smaller diameter tubes also result in a more compact boiler,
8-4
 M. A. Elmasri, 2000-2008
Chapter 8: Heat RecoveryBoilers
since the ratio of their weight to their surface area is smaller than the larger tubes. With smaller tubes, the
headers may also be of smaller diameter, and therefore tubes and headers are less thick, and although more
tubes will be used in total, the overall result is a lighter boiler with faster thermal response. For a given heat
transfer duty and pressure drop, a boiler made of 1½ " tubes would weigh about 20% less than one made of
2" tubes.
Tube wall thickness is selected based on pressure, but a minimum thickness is also needed for mechanical
integrity. For 2" (50.8mm) diameter tubes, some designs will use a minimum thickness of 0.105" (2.67
mm), which is thicker than needed from a pressure viewpoint below about 500 psi (35 bar). More
conservative designs will use a minimum thickness of 0.12" (3.05 mm), which is thicker than needed from a
pressure viewpoint below about 800 psi (55 bar).
1.c. Selection of Finned Tube Materials:
The fins and the tubes are normally made of the same material. Carbon Steel is chosen for most
economisers and evaporators, since it is strong and inexpensive and also has high thermal conductivity
compared to alloy steels. If the water is well deaerated, corrosion is negligible. The low-temperature
economiser, sometimes called the condensate pre-heater, which may carry cool condensate that has not yet
passed through the deaerator, should be made of Stainless Steel. This also helps it resist corrosion if water
were to condense from the flue gases on its outer surface. Even when burning sulphur-free natural gas, such
condensate contains dilute Carbonic acid.
The high-temperature superheaters and reheaters are usually made of alloy steels to better resist corrosion.
T11 (1¼Cr – ½Mo) or T22 (2¼Cr – 1Mo) are used when there is no supplementary firing, with tube metal
temperatures below about 1100 ºF (590 ºC). Their thermal conductivities are about 80% that of Carbon
Steel. T91 (9Cr – 1Mo) is stronger, and is used for higher pressures, but still limited temperatures, and its
conductivity is about 60% that of Carbon Steel. With high gas temperatures associated with supplementary
firing, a Stainless Steel may be used.
In high-pressure designs, the fins may be made of a more conductive material, and the tube of a stronger
material to avoid an overly thick tube wall. Also, the fins, which operate at higher temperatures, may be
made of a more corrosion-resistant material than the tube It is common to use stainless fins (typically
TP409) for high-temperature corrosion resistance, on T11, T22 or T91 tubes, which have good strength and
are less expensive than stainless. The thermal conductivity of stainless, however, is low (409 has a
conductivity about 60% that of Carbon Steel). Fin material has a much greater influence than tube material
on the overall thermal calculations.
1.d. Fouling Factors:
Water-side fouling factor is applied to the tube inside area, so has a greater influence on overall thermal
resistance than the gas-side factor, which is applied to the much larger outside area of a finned tube. Waterside fouling depends on deposition of scale, which, in turn, depends on water treatment quality and on
blowdown. With good boiler water and operating practice, water-side fouling factors should be in the range
of 0.0005 – 0.002 [BTU/hr-ft2-ºF]-1 , i.e. about 0.0001-0.00035 [W/m2-ºC]-1 depending on the age of the
boiler. Gas-side fouling depends on type of fuel, cleanliness of the air in the plant's vicinity, quality of the
GT's inlet air filtration, and corrosion resistance of the fin materials. Values in the range 0.001 – 0.01
[BTU/hr-ft2-ºF]-1 , i.e. about 0.0002-0.002 [W/m2-ºC]-1 can be expected, with the lower-end of that range for
a clean, well-filtered plant that only burns natural gas. With finned tubes, water/steam-side fouling has a
8-5
 M. A. Elmasri, 2000-2008
Chapter 8: Heat RecoveryBoilers
much greater influence on the overall heat transfer coefficient than gas-side fouling, because it acts on the
much smaller tube wall inside area.
2. BOILER DUCT
A good rule of thumb in estimating the cross section of an HRB duct is that the gas mass flux should be in
the range 0.65-1 lb/ft2-s (3-5 kg/m2-s). The lower end of this range applies to boilers with a small mean
temperature difference, such as with triple-pressure reheat cycles, and the upper end of the range applies to
boilers with a large mean temperature difference, such as with a single-pressure HRB making process steam.
For example, the HRB of a triple pressure reheat cycle based on a GE Frame 7F, which has an exhaust flow
of about 1000 lb/s (450 kg/s), would have a duct cross-section of about 1000/0.65 ≈ 1500 ft2 (450/3 ≈150
m2). Similarly, a single-pressure HRB making process steam from the exhaust of a GE LM2500+, which
has a flow of about 190 lb/s (85 kg/s), would have a duct cross-section of about 190/1 ≈ 190 ft2 (85/5 ≈ 18
m2).
A further simplification of the above rule of thumb, is to note that most modern gas turbines have a specific
power of about 170 kW per lb/s of exhaust gas. Thus, very crudely, the boiler duct cross-section will be
about 6-9 ft2 (0.55-0.85 m2) per MW of gas turbine output, with the lower end applying to basic
cogeneration HRSG's and the upper end to complex combined cycle HRSG's.
The duct Aspect Ratio, Tube Length / HRSG Duct Width, is typically in the range 1 to 3.5. Higher values
give a smaller number of long tubes, reducing manufacturing costs, particularly for smaller HRSG's. Large
boilers use lower values to avoid excessively-long tubes, which would be difficult to support against
vibration, and to handle and transport in pre-fabricated panels.
The exhaust gas leaving the last turbine wheel has an axial velocity of about 900 ft/s (275 m/s). The
engine's exhaust diffuser duct will slow it down by a factor of nearly two, to about 500 ft/s (150 m/s). Thus,
the heat recovery boiler duct cross-section is roughly ten times that of the gas turbine's exhaust diffuser, and
will need a connecting transition duct with an area expansion of roughly ten between the GT and the main
HRSG duct, which contains the tubes. To reduce pressure loss and avoid recirculating pockets of gas, this
expansion should not be too rapid, which means that a considerable amount of space may be consumed by
the transition duct.
The boiler gas-tight duct is usually made of Carbon Steel, about ¼" (6 mm) thick, with appropriate
stiffening structural steel sections welded at intervals. For very large HRSG's, thicker steel, such as 5/16" (8
mm) or even 3/8" (10 mm) may be used. The stiff gas-tight structure is usually insulated on the inside, with
a Stainless Steel liner protecting the insulation from the hot gases. That liner is typically made of
overlapping squares, each about 4' (1.2 m) square, with a few inches of overlap over it neighbours. Each
sheet metal square is supported by several stand-offs from the structural skin, with insulation between them.
The liner squares can slide relative to their neighbours as they expand or contract. The bottom floor liner
may be made of 10 gauge (0.135"/3.4mm) or 11 gauge (0.12"/3mm) stainless, since it may need to support
the weight of repair crews and their equipment. The sides and roof are typically made of thinner sheets, 12
gauge (0.105"/2.7mm) to 14 gauge (0.075"/1.9mm).
The typical boiler duct construction described above is normally used without supplementary firing, or with
mild levels of supplementary firing, up to about 1400 ºF (760 ºC). Although such designs are sometimes
quoted with firing up to 1600 ºF (870 ºC), the liner durability in withstanding temperature cycling up to that
level, for many years without warping, is questionable. Once the liner sheets begin to warp, the flowing gas
creates vortices which tend to make that situation worse, particularly if these vortices begin to entrain the
8-6
 M. A. Elmasri, 2000-2008
Chapter 8: Heat RecoveryBoilers
insulation, one particle at a time, until a significant amount of insulation is ripped out, causing further
problems by heat-stressing the outer, structural casing. If the HRSG has SCR, the insulation entrained with
the gas will clog it. Thus, it would be wise to specify either ceramic or refractory brick lining, or preferably
water walls, in the zones of the HRSG which will routinely cycle up in temperature with supplementary
firing to above 1400 ºF (760 ºC). Ceramic liner of the type shown in Fig. 12 does not cost much more than a
stainless steel liner, whereas water-wall lined duct may cost about $500/ft2 ($5000/m2), about four times a
lined duct.
1
2
FOR QUALITATIVE INDICATION ONLY
A
B
3
4
5
6
7
8
200 °F (93 °C)
50 ft/s (15 m/s)
Typical numbers, 3-PRH CC, 180 MW GT + 100 MW ST:
Weight ~ 1,750 tonnes
H.T. surface area ~ 135,000 m2 ~ 1,500,000 ft2
Cost ~ $16-$24 million, depending on DB, SCR, CO catalyst, bypass stack
Specific cost ~ $150/m2 ~ $13/ft2 ~ $11/kg ~ $5/lb ~ $200/ST MW
Gas mass flux: 0.65-1 lb/s per sq.ft. (3-5 kg/s per sq.m.)
Duct cross-section: 6-9 sq.ft. (0.6-0.9 sq.m.) per GT MW
Duct aspect ratio: 1-3.5
Overall mean ∆T: 75 ºF (42 ºC)
B
C
C
F
55 ft/s (17m/s)
1100 °F (590 °C)
500 ft/s (150 m/s)
D
A
24 ft/s (7m/s)
G
D
H
E
E
C
A
D
E
Thermoflow, Inc.
Company: TF
User: Maher Elmasri
HEAT RECOVERY STEAM GENERATOR
F
A
B
33.5 ft
1
C
D
E
46.9 ft
57.2 ft
13.4 ft
2
F
118 ft
3
G
H
I
J
64.4 ft
17.9 ft
-
-
4
5
ELEVATION
Date: 11/25/00
F
Drawing No:
c:\TFLOW4\MYFILES\GTPRO.GTP
6
7
8
PEACE/GT PRO 10.0.1
Figure 3. Horizontal HRB
Figure 3 shows a "Horizontal Boiler", in which the exhaust gas flows horizontally through the main HRSG
duct, which contains vertical tubes perpendicular to the gas flow. Figure 4 shows a "Vertical Boiler", in
which the exhaust gas flows vertically upwards through the main HRSG duct, across tubes laid out
horizontally, within the duct. Although early American designs in the 1970's were of the Vertical type,
virtually all of the current American manufacturers have adopted the Horizontal design. By contrast, several
of the major European manufacturers have preferred the Vertical design.
Many claims have been made regarding the inherent superiority of one type over the other. The opinion of
this writer is that in many circumstances, the differences between detailed design practices of the various
vendors, for either type, matter much more than the orientation of the gas flow. Further complicating the
horizontal vs. vertical debate is the fact that most vendors of vertical boilers have historically employed
forced (or "assisted") circulation, whereas most manufacturers of horizontal boilers have historically used
natural circulation. The fact is that either mode of circulation can be employed in either design, and
preference for one mode of circulation or another should not be confused with preference for gas flow
direction.
8-7
 M. A. Elmasri, 2000-2008
Chapter 8: Heat RecoveryBoilers
1
2
FOR QUALITATIVE INDICATION ONLY
3
4
5
6
7
8
A
A
H
B
B
G
C
C
F
D
D
E
I
E
A
C
E
D
Thermoflow, Inc.
Company: TF
User: Maher Elmasri
HEAT RECOVERY STEAM GENERATOR
F
A
B
16.1 ft
1
C
D
E
F
G
H
17.2 ft
62.0 ft
34.4 ft
55.1 ft
15.5 ft
52.5 ft
2
3
I
17.2 ft
4
ELEVATION
J
Date: 11/25/00
5
F
Drawing No:
c:\TFLOW4\MYFILES\GTPRO.GTP
6
7
8
PEACE/GT PRO 10.0.1
Figure 4. Vertical HRB
Nevertheless, it may be convenient to summarise the views expressed by the proponents of each type, with a
comment on each issue.
Claims made for the superiority of horizontal boilers are:
1. That natural circulation is more reliable, since it does not depend on an auxiliary pump, and it also
saves the cost of the pump (and its backup unit) and its auxiliary power consumption, not to mention
having one more critical item to maintain. Since many vertical boilers built nowadays do operate with
natural circulation, this issue is not as relevant as it once was, when vertical boilers were synonymous
with forced circulation. Recent designs of vertical assisted-circulation boilers use a small pump to start
the circulation, then shut it down when the boiler is operating. The pump only operates on a small
percentage of the flow, injecting it through a jet into the main flow, thereby inducing it to flow as well.
When the pump is shut-off, the pipes leading to it and from it are valved off, and the boiler circulates
naturally.
2. Easier to erect in the field and to modify, if needed. For example, additional sections of empty duct
can be installed initially, and may be used to conveniently add heat transfer surface, or SCR, etc. There
is some merit in this argument regarding modifications, but not regarding ease of erection, which
depends more on the extent of prefabrication and modularity of the boiler.
8-8
 M. A. Elmasri, 2000-2008
Chapter 8: Heat RecoveryBoilers
3. Easier to install and to replace SCR or CO catalyst. This argument is of dubious merit, since once
provision is made for SCR in a vertical boiler, sliding the spent catalyst panels out and the new ones in
should be quite feasible.
4. Rain, or water from soot-blowing or pressure-washing of the tubes would not fall down into a
horizontal catalyst and clog it. This is a valid point in the uncommon cases where the GT burns dirty
fuel, but boilers working on exhaust from natural gas fired turbines do not require soot blowing or
power washing, and any rainwater or minor debris that find their way to the SCR when the GT is shut
down will be blown away when the GT starts.
Claims advanced for the superiority of vertical boilers are:
1. Forced (or "assisted") circulation provides faster startup and better transient response. This claim
has merit, but noting that many vertical boilers built nowadays operate with natural circulation, this
issue is more a function of circulation method than orientation. In other words, it is possible to build
horizontal boilers with assisting pumps or water jets to help initiate faster circulation. A factor which
has clouded this argument is the fact that major European manufacturers of vertical boilers have tended
to utilise smaller-diameter tubes, which result in a lighter boiler with faster warm-up and transient
response. Using smaller tubes with assisted circulation in horizontal boilers would provide these
benefits to the horizontal type as well.
2. Less need for Real Estate. This has some merit, in special cases where a plant has to be built into a
very restricted site. In most cases, however, anyone who can afford to build a power plant should be
able to afford purchase of an extra half-acre of land ! It may also be noted that saving in plot plan area
is not very much at all, if the gas turbine has an axial exhaust, which most modern machines do. The
older heavy duty gas turbines with a side exhaust (see Fig. 4 in the chapter on "Gas Turbine
Components – Combustors") could be placed under the boiler., exhausting straight up into it, saving a
substantial area of land, but at the expense of complicating the installation and maintenance of the gas
turbine and boiler.
In summary, the present writer sees no clear advantage of one type or another. Differences in design
philosophy between manufacturers, with regard to using assisted circulation or relying on purely natural
circulation, and with regard to tube diameters, are more important factors.
3. POLLUTION CONTROL CATALYSTS IN THE BOILER
3.a. Selective Catalytic Reduction (SCR)
The exhaust of a typical gas turbine burning natural gas, without Dry Low NOx (DLN) combustors may
contain 150-250 ppm of Nitrogen Oxides (NOx). Using steam or water injection can reduce that to 25-45
ppm, and recently developed DLN combustors can achieve about 25 ppm withour water or steam injection.
However, in many regions of the USA, Japan, and a growing list of countries, environmental regulations
require even lower NOx levels, in some cases below 10 ppm. To achieve them, Selective Catalytic
Reduction (SCR) has been developed.
SCR uses ammonia as a reducing agent, to take the oxygen out of the NOx, in the presence of oxygen,
according to the overall reaction, in which NOx is treated as NO :
NO + NH3 + ¼ O2 → N2 + 1½ H2O ………. (1)
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 M. A. Elmasri, 2000-2008
Chapter 8: Heat RecoveryBoilers
This overall reaction cannot take place directly, but is the net result of a series of sub-reactions that take
place on the surface of a catalyst, within a narrow range of temperatures of 625-825 ºF (330-440 ºC). Thus,
the SCR is frequently placed behind the HP superheater but before the HP evaporator in the gas path, or in
between the rows of HP evaporator tubes. The exhaust gases from open-cycle gas turbines, such as peakers,
is too hot for SCR. There have been applications where water was injected into the exhaust to cool it down,
just to be able to use SCR on a peaker.
Figure 5a (Left) SCR catalyst in a horizontal HRSG.
Figure 5b (above) Ammonia injection skid
(courtesy of MassPower)
The catalyst is either Platinum, or Base Metal Oxides, Vanadium Pentoxide (V2O5), or Titanium Oxide
(TiO2). These are either deposited on a porous ceramic "sponge" or a mesh of stainless steel foil coated with
Alumina. It should present a large surface area of contact with the gas, with minimal pressure drop. Typical
constructions provide a surface/volume ratio of about 50 in2/in3 (2 mm2/mm3) and an open flow area of
about 80%. Gas-side pressure drop of 1-2 "H2O (2.5-5 mbars) are typical, depending on the required
effectiveness. Figure 5 shows an SCR catalyst within a horizontal boiler.
As equation (1) shows, the ammonia mole flow rate should be equal to the mole flow rate of NO, so noting
that the molecular weight of NO is 30 and that of NH3 is 17, we find that for each lb or kg of NO, 17/30 =
0.57 lbs or kgs of ammonia are needed. Injecting more ammonia results in ammonia slip into the
atmosphere, while injecting less means some NOx will escape. In addition to the correct amount of
injection, the ammonia should be well-mixed with the gas before passing through the catalyst, and is
therefore vapourised and mixed with air, or steam, prior to injection from a grid upstream of the catalyst.
Figure 6 shows an ammonia injection skid, where it is vapourised and blown into the HRSG grid by fans. If
the aqueous ammonia is vapourised using electric heaters, these heaters would consume an auxiliary load of
about 0.1-0.15% of the GT output, and the fans about 0.015%-0.02% of the GT power. If the aqueous
ammonia is vapourized using hot gas extracted from the HRSG, the electric heater load is saved, but the fans
require more power, about 0.04% of GT power, and the system is more complicated.
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 M. A. Elmasri, 2000-2008
Chapter 8: Heat RecoveryBoilers
Assuming we have 25 ppm of NO to begin with, we would need to inject 14 ppm of ammonia. Thus, a large
gas turbine such as a Frame 7F, with an exhaust flow of 1000 lb/s, would need 0.014 lb/s of ammonia,
which is about 50 lbs/hr (22 kg/hr) or roughly half a ton per day or 3½ tons per week. Since ammonia is
trucked and stored as aqueous solution, this corresponds to a large tanker truck-load every week. Ammonia
is a poisonous substance in any significant concentration, so the danger of such a hazardous substance being
trucked around in large amounts poses collateral environmental dangers.
In theory, the SCR catalyst should last indefinitely, but in practice it gets poisoned or eroded and must be
replaced. Typical life is about five years.
SCR was developed in the early 1980's, and despite its various drawbacks, is now an accepted standard, at
least in the USA. As with mature technologies, the cost has come down (in real terms) since the 1980's. For
a 200-MW-class GT (300-MW-class CC), the additional equipment costs ~ $3 million, and fully installed
adds ~ $6 million.
3.b. The SCONOX Process
This is a new process that eliminates the environmental risks of ammonia. Many panels of noble metal
catalyst coated with Potassium Carbonate K2CO3. are placed in the HRSG duct, in a location where the gas
temperature is 300-700 ºF (150-370 ºC). In the presence of the catalyst, a reaction takes place in which the
NOx is absorbed, changing the coating on the catalyst to Potassium Nitrite KNO2 and Potassium Nitrate
KNO3. After about 15 minutes, the catalyst panel is isolated by louvers, and a gaseous mixture of H2 and
CO2 introduced to regenerate it, converting the Potassium Nitrate and Nitrite back to Potassium Carbonate,
and releasing the Nitrogen as N2. This takes 3-5 minutes. Thus, at any given moment, about ¾ of the
catalyst panels are open to the exhaust flow, absorbing NOx, and ¼ are shut-off by louvers, undergoing
regeneration.
This system has been reported as functioning effectively since 1997, lowering NOx to 3.5 ppm in LM2500
pilot installations in California. (Diesel & Gas Turbine Worldwide, Jan/Feb 1998).
Figure 6a illustrates the system layout and Figure 6b shows an actual installation.
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 M. A. Elmasri, 2000-2008
Chapter 8: Heat RecoveryBoilers
Figure 6a. Schematic of SCONOX system (Courtesy of EmeraChem Corp.)
Full View of Plant
With EMx GT in Operation
Actuator Arms
For Opening and
Closing Chambers
Regeneration Gas
Delivery System
Figure 6b. Photos of a SCONOX istallation (Courtesy of EmeraChem Corp.)
3.c. CO Catalytic Oxidation
CO oxidation is simpler than NOx reduction because all the ingredients of the reaction
CO + ½ O2 → CO2 ………. (2)
are already present in the exhaust gases. The catalyst just serves to accelerate a reaction which would
normally occur anyway, had the concentration of CO and the residence time been sufficient. The CO
catalyst can be used in a wider range of temperatures, and may even be installed directly in the GT's exhaust.
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 M. A. Elmasri, 2000-2008
Chapter 8: Heat RecoveryBoilers
Figures 7a & 7b. Replacement tube panels for a horizontal boiler
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 M. A. Elmasri, 2000-2008
Chapter 8: Heat RecoveryBoilers
Figures 8a & 8b. Replacement tube panels in a horizontal boiler
8-14
 M. A. Elmasri, 2000-2008
Chapter 8: Heat RecoveryBoilers
Figures 9a & 9b. Shop assembly of
tube/header modules for a horizontal
boiler (courtesy of ALSTOM Boilers)
8-15
 M. A. Elmasri, 2000-2008
Chapter 8: Heat RecoveryBoilers
Figures 10a & 10b. Lifting a shop-assembled module of a horizontal boiler for field installation
(courtesy of ALSTOM Boilers)
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 M. A. Elmasri, 2000-2008
Chapter 8: Heat RecoveryBoilers
Figures 11a & 11b. Site erection of a horizontal boiler (courtesy of ALSTOM Boilers)
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 M. A. Elmasri, 2000-2008
Chapter 8: Heat RecoveryBoilers
Figure 12. Ceramic liner installed in a horizontal boiler, downstream of a duct burner (courtesy of
ALSTOM Boilers)
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 M. A. Elmasri, 2000-2008
Chapter 8: Heat RecoveryBoilers
Figure 13. Vertical boiler with by-pass stack (courtesy CMI)
Figure 14. Frame for a large, modularly erected vertical boiler (courtesy of CMI)
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 M. A. Elmasri, 2000-2008
Chapter 8: Heat RecoveryBoilers
Figure 14 (contd.). Modular erection of a vertical boiler (courtesy of CMI)
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 M. A. Elmasri, 2000-2008
Chapter 8: Heat RecoveryBoilers
Figure 14 (contd.). Modular erection of a vertical boiler (courtesy of CMI)
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 M. A. Elmasri, 2000-2008
Chapter 8: Heat RecoveryBoilers
Figure 14 (contd.). Modular erection of a vertical boiler (courtesy of CMI)
8-22
 M. A. Elmasri, 1990-2002
Chapter 9: Single-Pressure HRB/CC Thermodynamics
THERMODYNAMICS OF SINGLE-PRESSURE HEAT
RECOVERY BOILERS & COMBINED CYCLES
Content Revised May, 2002
© Maher Elmasri 1990-2002
Single-pressure HRB’s are widely used in cogeneration plants to generate process steam. They are also
commonly used in the simpler combined cycles arrangements. Most large combined cycles use multipressure HRB’s. Because of the simplicity of the single-pressure system, it is discussed first to provide
insight into heat recovery and combined cycle design principles.
The results shown here are all from the design perspective, meaning that thermodynamic parameters are
selected and equipment sized to satisfy them. Off-design behaviour of fixed equipment with under varying
conditions is discussed elsewhere. Interestingly, it will be seen that the trends presented here for designpoint calculations are very similar to those at off-design for fixed-equipment calculations.
Typical gas turbines have exhaust temperatures in the range 430-510 °C (806-950 °F) for aeroderivatives
and 530-610 °C (986-1130 °F) for heavy duty engines. The exhaust gases carry roughly 60% of the fuel’s
LHV energy. Hypothetically, all this energy (as well as some additional latent heat) can be recovered if the
gases were cooled down to the ambient temperature in a heat recovery boiler. This cannot be accomplished
since there is a minimum practical stack temperature. In addition, fundamental thermodynamic constraints
may dictate a stack temperature higher than the practical minimum.
1. PRACTICAL CONSTRAINTS ON MINIMUM STACK TEMPERATURE
1.a. Plume Dispersion Constraint:
The exhaust gases leaving from the boiler stack need to be hot enough to be buoyant, so they can rise higher
into the atmosphere as they mix with air, to avoid causing high levels of pollutant concentrations in the
vicinity of the plant. Naturally, if the pollutants are relatively benign, such as from efficiently burning
sulphur-free natural gas in low-NOx systems, the pollutant dispersion problem is not as restrictive as with
plants that emit larger concentrations of unburnt hydrocarbons (UHC), nitrogen oxides (NOx), sulphur
oxides (SOx) and particulates. Besides the pollutant dispersion issue, condensation of the water vapour in
the flue gases produces an unsightly plume, which though benign, may be disliked by the populace. Hotter
flue gas discharge reduces the extent of the condensation plume.
The pollutant dispersion problem can be ameliorated by building a taller stack and by discharging the gases
at a higher temperature and velocity. A tall stack is expensive as well as aesthetically displeasing and
hazardous to aircraft. High gas discharge velocity results in a higher back-pressure on the gas turbine.
Depending on the site’s environmental restrictions and permitting authorities, detailed studies of plume
dispersion and pollutant concentrations may need to be submitted. These studies consider stack discharge
temperature and velocity, as well as wind and weather patterns at the site and are usually performed by
specialised consultants. The hotter the stack temperature, the easier it becomes to satisfy the pollutant
dispersion requirements with a reasonable stack height and velocity.
These issues are very much site-specific, but as a crude rule of thumb, stack temperatures below about 85 °C
(185 °F) are generally unfeasible from a dispersion viewpoint, even with the cleanest natural gas burning
systems. With other fuels, or with environmentally sensitive sites, the practical stack temperature from a
dispersion point of view is higher, typically in the 95 °C (205 °F) range for plants burning natural gas.
9-1
 M. A. Elmasri, 1990-2002
Chapter 9: Single-Pressure HRB/CC Thermodynamics
1.b. Water Vapour Condensation Constraint:
The exhaust gases from a gas turbine burning CH4 have a water vapour content of 7-10% by volume,
resulting in a H2O dewpoint in the range 40-45 °C (104-113 °F).
The H2O in the flue gas is increased by fuel burnt in supplementary firing. If natural gas is used in the duct
burner, each 100 °C (180 °F) of duct burner temperature rise corresponds to an increase in flue gas H2O
content of roughly 0.8% by volume, so that in a supplementary fired HRB, the water vapour content in the
flue gas may reach 12% by volume, with a corresponding H2O dewpoint of about 50 °C (122 °F). The water
vapour content in the flue gases is further increased if gas turbine steam injection is employed. In the
(uncommon) case of massively steam-injected gas turbines, it may reach 25% by volume, with a
corresponding dewpoint of about 65 °C (149 °F).
To avoid condensation of H2O from the flue gases, the coldest tube-surface metal temperature within the
boiler should be above the dewpoint. In a typical finned-tube economiser, depending on fin and tube
materials and fouling factors, the minimum metal temperature is above the feedwater temperature by 1525% of the difference between the flue gas and feedwater temperatures, as illustrated in Fig. 1 below.
Tm = Tfw + x (Tg – Tfw)
Tm
..................................
Tg
(1)
Tm = Tfw + x (Tg – Tfw)
For a typical economiser:
x ≈ 0.15-0.25
Tfw
Figure 1. Temperatures in a Finned-Tube Economiser
Thus, the stack temperature at which condensation would occur is when the metal reaches the dewpoint:
Tg, condensation = Tfw + (Tdewpoint – Tfw) / x ................ (2)
For example, if the feedwater in a combined cycle is at 40 °C (104 °F), and the flue gas dewpoint is 50 °C
(122 °F), assuming that x=0.2, we can estimate that condensation on the economiser would occur if the stack
temperature were = 40 + (50 – 40) / 0.2 = 90 °C (194 °F). This sets a minimum stack temperature to avoid
condensation.
In practice, warm feedwater from the economiser’s exit can be recirculated, via a pump, back to its inlet to
raise Tfw as needed to prevent condensation. The rate of recirculation is controlled by the measured stack
temperature, and is automatically increased if the stack temperature falls. If the feedwater within the tubes
were heated to the dewpoint - 50 °C (122 °F) in the above example - the stack temperature could be lowered
to any desired level above 50 °C (122 °F) without possibility of condensation.
Although the vast majority of practical boilers are designed and operated to avoid flue gas H2O
condensation, a few installations have been designed with condensation, to recover additional heat, sensible
and latent, as well as to recover water. One cogeneration facility with a massively steam-injected gas
turbine has been installed in Italy in a car body manufacturer’s paint shop. The low temperature heat
9-2
 M. A. Elmasri, 1990-2002
Chapter 9: Single-Pressure HRB/CC Thermodynamics
recovery coils produce hot water, while cooling the flue gases to very low stack temperature of about 65 °C,
thereby condensing part of their water vapour content. The finned tube coils over which the condensation
occurs are made of plastic, and laid out horizontally, with gutters to collect the condensate. Overall
efficiency based on LHV may exceed 100% when some of the combustion-generated H2O is condensed, in
addition to all the steam injected into the GT.
1.c. Acid Dewpoint Constraint:
Natural gas may contain hydrogen sulphide (H2S) or sulphur dioxide (SO2). After the gas is treated, if
necessary, sulphur content is usually below 0.001% (10 ppm) by weight. Distillate oil used in gas turbines
typically contains 0.1-0.2% sulphur by weight, and most GT manufacturers will specify a maximum sulphur
content in the fuel of 0.5% by weight.
The sulphur in the fuel is converted to SO2 during combustion. According to
S + O2 → SO2,
each kg of S (Atomic weight=32) produces 2 kgs of SO2 (sulphur dioxide- molecular weight=64).
A modern, high specific power gas turbine burning distillate oil has an air/fuel ratio of about 40:1 by weight
(about 45:1 by weight if burning gas). Thus, if a fuel oil contained 0.1% S by weight, the S concentration in
the exhaust gases would be about 1 part in 40,000 (25 ppm) by weight; and the SO2 concentration would be
about 1 part in 20,000 (50 ppm) by weight.
Although noxious, SO2 is not corrosive. However, some SO2 is oxidized to produce SO3, which has a great
affinity for H2O, with which it reacts to form the highly corrosive H2SO4 (sulphuric acid),
SO2 + ½ O2 → SO3
SO3 + H2O → H2SO4 .
The percentage of SO2 converted to SO3 depends on the available oxygen concentration and the timetemperature history of the flue gases. Based on empirical data, it is generally assumed at 1-2% for
conventional fired boilers, but higher for gas turbines, at up to 5%. With a 5% conversion rate, an oil fuel
with 0.1% S will result in a GT exhaust with about 3 ppm SO3 by weight.
The acid dewpoint, the temperature at which sulphuric acid will condense, depends on the partial pressure of
SO3 and H2O in the flue gas. A relationship is given by Pierce in Reference [1] •.
Figure 2 shows acid dewpoint as a function of fuel sulphur content, computed in GT PRO, which uses the
method given in Reference [1]. The figure is for a typical distillate oil. Two types of gas turbines are
considered: a modern GT with a high specific power of 400 kW per kg/s airflow (180 kW per lb/s); and an
older GT with moderate specific power of 250 kW per kg/s airflow (115 kW per lb/s). The modern GT has
a lower excess air, hence a higher concentration of SO2 in its flue gases, resulting in a higher acid dewpoint.
Two curves for each GT are provided, one assuming 5% conversion of SO2 to SO3 (GT PRO’s default,
thought to be conservative) and one assuming just 2% conversion, thought to be the lower likely bound.
Figure 2 is for GT exhaust, without supplementary firing. If a duct burner is used, it will increase the flue
gas H2O content, which raises the acid dewpoint even if no additional sulphur is introduced with the duct
burner fuel. If the duct burner fuel contains sulphur, it will further raise the acid dewpoint.
•
Pierce, Robert R., “Estimating Acid Dewpoints”, Chemical Engineering, April, 1977.
9-3
 M. A. Elmasri, 1990-2002
Chapter 9: Single-Pressure HRB/CC Thermodynamics
A natural gas fuel will result in a slightly higher acid dewpoint than a distillate oil if both have the same
sulphur content by weight. Generally, however, natural gas has a far lower sulphur content than distillate
oil.
The method to compute the minimum stack temperature as a function of feedwater temperature to avoid acid
condensation is similar to that given in §1.b above for H2O condensation, except that the acid dewpoint
should be used rather than the H2O dewpoint.
Estimated Sulphuric Acid Dewpoint as a Function of Distillate Oil Sulphur Content
130
266 F
125
248 F
Acid Dewpoint, C
120
High Specific Power GT, 5%
SO2 conversion
115
230 F
110
Low Specific Power GT, 5%
SO2 conversion
High Specific Power GT, 2%
SO2 conversion
Low Specific Power GT, 2%
SO2 conversion
105
212 F
100
95
90
0
0.1
0.2
0.3
0.4
0.5
194 F
Fuel Sulphur Content, % by weight
Figure 2. Estimated Flue Gas Acid Dewpoints with Various Assumptions
Many HRB’s are designed to operate primarily with natural gas, and with distillate oil as a backup fuel.
Thus, an efficient design would recover heat down to a fairly low stack temperature, such as 90 °C (194 °F),
if possible. When oil is used instead of gas, the stack temperature would be raised to prevent acid
condensation. This may be done by either recirculating warm water from the economiser exit back to its
inlet, or by by-passing cool water around the economiser. The former method requires an additional pump,
but allows one to avoid acid condensation without raising the stack temperature as much as the latter method
would necessitate.
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 M. A. Elmasri, 1990-2002
Chapter 9: Single-Pressure HRB/CC Thermodynamics
1.d. Overall HRB Energy Balance, from Inlet to Stack:
Neglecting minor heat losses and blowdown, the energy balance around the boiler relates the steam
generation rate relative to the gas flow rate as follows:
Exhaust gas out at stack
Exhaust gas in
m g hg , stk
m g hg ,1
Heat Recovery Boiler
Feedwater in
Steam out
ms hs
ms h fw
Figure 3. Energy Balance (First Law) Applied to the Entire HRB
ms (hs − h fw ) = m g (hg ,1 − hg , stk )
where
 (hg ,1 − hg , stk ) 
∴ ms = m g 

 (hs − h fw ) 
..................................... (3)
............................................. (4)
m g is the flue gas mass flow rate
hg ,1 is the gas specific enthalpy at the inlet to the HRB, evaluated at the inlet temperature Tg ,1
hg , stk is the gas specific enthalpy at the stack, evaluated at the stack temperature Tg , stk
ms is the steam mass flow rate
hs is the final steam specific enthalpy leaving the HRB superheater
h fw is the initial feedwater specific enthalpy entering the economiser
2. PINCH CONSTRAINT ON MINIMUM STACK TEMPERATURE
Equation (4) can be used to find the steam generation rate if the stack temperature were known or could be
correctly assumed a priori. In addition to the practical constraints on the minimum stack temperature given
in §1.a-§1.c, the thermodynamics of the temperature profile within the HRB may further constrain the
minimum stack temperature, as illustrated below.
2.a. Simplistic Application of the Overall Energy Balance with an Assumed Stack Temperature:
Let us assume that we wish to recover heat from the 550 °C (1022 °F) exhaust gases of a typical, mediumsized gas turbine. The exhaust heat will be used to generate steam at 50 bar/530 °C (725 psia /986 °F) for a
steam turbine in a combined cycle. The feedwater, after being pumped from the condenser, is at 34 °C (93
°F)/65 bar (943 psia). Suppose that, as in many actual plants burning natural gas, the practical constraints
would allow a minimum stack temperature of 90 °C (194 °F), which we assume that we can indeed achieve.
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 M. A. Elmasri, 1990-2002
Chapter 9: Single-Pressure HRB/CC Thermodynamics
We apply Equation (4), assuming (naively) that Tg , stk = Tstk , min =90 °C (194 °F). We find the gas enthalpy
drop from 550 to 90 °C (1022 to 194 °F) as (hg ,1 − hg , stk ) =506 kJ/kg (217.5 BTU/lb) of gas [assuming flue
gas composition is 75.8% N2Ar, 14% O2, 3.1% CO2 and 7.1% H2O]. From the water/steam properties we
find the enthalpy difference between steam and feedwater as (hs − h fw ) =3355 kJ/kg (1442 BTU/lb) of
steam. Substituting this property data in Equation (4), we calculate that ms = 0.1508m g , i.e. the steam
generation rate is 15.08% of the flue gas flow rate.
Exhaust gas out, assumed @ 90 °C
Exhaust gas in @ 550 °C
m g hg ,1
m g hg , stk
Heat Recovery Boiler
Feedwater in @ 65 bar/34 °C
Steam out @ 50 bar/530 °C
ms hs
ms h fw
The calculation given above is based on conservation of energy, without consideration of the temperature of
the heat source vis a vis that required by the heat sink, except at the boundaries. The flue gases are hotter
than the steam by 20 °C (36 °F) at the hot-end and by 56 °C (101 °F) at the cold-end. How about the
temperature distribution within the HRB? Is the gas hotter than the water/steam throughout the heat
recovery process? The Second Law of Thermodynamics requires that each kJ (BTU) of energy transferred
is from a source hotter than the sink. To answer this question, we plot the temperature profile of the gases as
they cool from 550 to 90 °C, and the temperature profile of the water/steam as it heats up from 34 to 530 °C.
Temperature [°C]
600
500
Flue Gas
400
Superheater
300
Evaporator
200
Economiser
100
0
0
10000
20000
30000
40000
Heat Transfer [kW]
50000
60000
Figure 4. T-Q Diagram Corresponding to Simplistic, Unachievable Heat Balance
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 M. A. Elmasri, 1990-2002
Chapter 9: Single-Pressure HRB/CC Thermodynamics
Fig. 4 shows the T-Q (Temperature vs. Amount of Heat Transfer) profile for superheated steam being
generated from subcooled water by heat from flue gases. The profile of flue gas cooling is essentially linear
on the T-Q diagram, since the gases have a nearly-constant specific heat. On the water-side, however, the
T-Q profile is comprised of three, near-linear sections with sharp breaks in their slope. The feedwater is
first heated in an economiser to saturation temperature (or just below it in typical practice). Next, the water
is evaporated at constant (saturation) temperature. Finally, the resulting steam is superheated to the final
temperature. These profiles are plotted for our numerical example, with a water/steam flow of 15.08% of
the flue gas flow, as calculated naively above (100 kg/s gas flow and 15.08 kg/s steam flow). Although the
gas is hotter than the water/steam at either end, due to the mismatch in the shapes of the temperature
profiles, such an energy balance would imply that the gas would need to transfer some of its heat to the
water/steam while it is colder than the water/steam. This is impossible. Thus the assumption that the flue
gas can be cooled to the desired temperature of 90 °C is unachievable.
2.b. Pinch-Constrained Heat Recovery:
The flue gas has to be hotter than the water/steam, at all points within the HRB, by a non-zero temperature
difference. The point along the heat recovery profile at which the minimum temperature difference occurs is
called the “pinch point”, and the minimum temperature difference is correctly called the “pinch point
temperature difference”, but we shall frequently refer to that temperature difference as the “pinch” for
brevity. As the temperature difference across which heat is transferred approaches zero, the necessary
surface area of the heat exchanger grows exponentially towards infinity. Thus, the minimum pinch cannot
approach zero. Its practical value is an economic consideration, discussed elsewhere. Values of the
minimum pinch between 10 and 20 °C (18 and 36 °F) are typical in good HRB design practice.
Assuming a minimum pinch of 15 °C (27 °F), our numerical example would yield the realistic temperature
profile shown in Fig. 5. The shape of the temperature profiles and the minimum pinch constraint conspire to
force a stack temperature of about 160 °C (320 °F). The actual heat recovered is only 433 kJ/kg (186
BTU/lb) of flue gas, and the steam production is only 12.9% of the flue gas flow. Thus, the pinchconstrained recoverable heat is less than calculated with the naive assumption of a 90 °C (194 °F) stack.
Figure 5. T-Q Diagram
for Realistic,
Heat Recovery
OTB element
[29] -Pinch-constrained
TQ Diagram
600
500
Pinch ∆T assumed
@ 15 °C (27 °F)
400
T [°C]
300
200
Min Tstk of 90 °C
(194 °F) cannot be
achieved
100
0
0
10000
20000
30000
Heat Transfer [kW]
9-7
40000
50000
 M. A. Elmasri, 1990-2002
Chapter 9: Single-Pressure HRB/CC Thermodynamics
With pinch-constrained heat recovery, Equation 4 still applies, except that the stack temperature is not
known a priori. The gas available enthalpy drop to the pinch-point governs the steam production rate, i.e.
 (hg ,1 − hg , pp ) 
ms = m g 

 (hs − hee ) 
............................................. (5)
Once the steam production rate is computed by Equation (5), the gas enthalpy at the stack may then be
found, by an energy balance over the economiser section, in which all quantities are now known except
hg , stk , i.e.
ms (hee − h fw ) = m g (hg , pp − hg , stk )
where
............................... (6)
hg , pp is the gas specific enthalpy at the pinch point, evaluated at the gas temperature at the pinch,
which
exceeds
saturation
Tg , pp = Tsat + ∆T pinch , min
temperature
by
the
pinch
temperature
difference,
i.e.
hee is the water specific enthalpy at the economiser exit (entering the evaporator)
Finally, the stack temperature Tg , stk is found from the stack specific enthalpy hg , stk .
2.c. Determining if Governing Constraint is Minimum Stack or Minimum Pinch:
To calculate how much steam can be generated in a single-pressure boiler, we need to decide on an
acceptable, minimum stack temperature Tstk , min and on an acceptable minimum pinch temperature
difference ∆T pinch , min . The steam generation rate is the lesser of that limited by the minimum stack and that
limited by the minimum pinch; i.e.
 (hg ,1 − hg , stk , min ) 
ms ≤ m g 

 (hs − h fw ) 
............................................. (7)
 (hg ,1 − hg @ Tsat + ∆Tpinch , min ) 
ms ≤ m g 

(hs − hee )


............................... (8)
These relations are already somewhat simplified, because minor heat losses and boiler blowdown have been
neglected. They may be further simplified by assuming that the constant-pressure specific heat of the flue
gases is constant, allowing the flue gas enthalpy differences to be expressed as the directly proportional
temperature differences,
(hg ,1 − hg , stk , min ) = C pg (Tg ,1 − Tstk , min )
............................... (9)
(hg ,1 − hg @ Tsat + ∆Tpinch , min ) = C pg (Tg ,1 − Tsat − ∆T pinch , min )
(10)
For typical GT exhaust gases, C pg is a function of temperature, varying from about 1.05 kJ/kg-°C (0.25
BTU/lb-°F) at 120 °C (250 °F) to about 1.175 kJ/kg-°C (0.28 BTU/lb-°F) at 600 °C (1112 °F). Its mean
9-8
 M. A. Elmasri, 1990-2002
Chapter 9: Single-Pressure HRB/CC Thermodynamics
value over a typical HRB temperature range is about C pg ≈ 1.1 kJ/kg-°C (0.265 BTU/lb-°F). C pg depends
on exhaust gas composition, increasing with H2O content.
*
Combining Equations (7)-(10), we can solve for the inlet gas temperature Tg ,1 that simultaneously satisfies
the minimum stack and the minimum pinch constraints:
*
Tg ,1 = Tg , p +
φ
1−φ
(Tg , p − Tstk , min )
C pg , pinch → stack
C pg ,1→ pinch
(11)
where Tg , p = Tsat + ∆T pinch , min is the flue gas temperature at the minimum pinch and φ is the ratio of heat
needed by the water/steam above the pinch (evaporator and superheater) to the total heat needed by the
water/steam. It depends only on the steam/water side enthalpies:
φ=
hs − hee
hs − h fw
............................. (12)
and where C pg , pinch → stack and C pg ,1→ pinch are the mean flue gas specific heats, averaged between pinch and
stack and between inlet and pinch, respectively.
Flue Gas Inlet Temperature for Transition between Minimum Stack and Minimum Pinch Constraints
for Single-Pressure Heat Recovery
365
725
1090
1450 psia
1815
2175
750
1382 F
Tstk,min=90 C, Tfw=34 C ,
Ts=550 C
Transition Gas Temperature (C)
Minimum Stack Temperature Limit
700
1292
Tstk,min=90 C, Tfw=50 C,
Ts=400 C
Tstk,min=120 C, Tfw=34 C,
Ts=550 C
650
1202
600
1112
Minimum Pinch Limit
550
1022 F
25
50
75
100
125
Tstk,min=120 C, Tfw=50 C,
Ts=400 C
90 C = 194 F
120 C = 248 F
34 C = 93 F
50 C = 122 F
550 C = 1022 F
400 C = 752 F
150
Steam Delivery Pressure (bar)
Figure 6. Flue Gas Transition Temperature for Single-Pressure Heat Recovery
*
If the gas entering the boiler is hotter than Tg ,1 , the recoverable heat is governed by the minimum stack
*
constraint, if the gas entering the boiler is cooler than Tg ,1 , the recoverable heat is governed by the
minimum pinch constraint. Thus, we call this inlet gas temperature the transition. Fig. 6 shows its values
across a range of pressures for the range of feedwater and steam temperatures typically associated with
combined cycles. These values are higher than typical gas turbine exhaust temperatures, so that without
supplementary firing, the minimum pinch governs the design of single-pressure heat recovery boilers.
9-9
 M. A. Elmasri, 1990-2002
Chapter 9: Single-Pressure HRB/CC Thermodynamics
3. EFFECT OF INLET GAS TEMPERATURE & STEAM PRESSURE
3.a. Effect of Inlet Gas Temperature on Pinch-Constrained Heat Recovery:
Figure 7 illustrates three T-Q% diagrams, all for generating steam at the same conditions of 50 bar/430 °C
(725 psia/806 °F), but with flue gases at three different inlet temperatures, 550, 600 and 650 °C (1022, 1112,
and 1202 °F). It is assumed that the minimum stack temperature is 90 °C (194 °F), so that all three cases are
pinch limited, since the inlet gas is below the transition temperature at the selected conditions. The figure
shows that when a single-pressure boiler is designed to produce steam at given conditions from feedwater at
a given condition, and with a given pinch-point temperature difference as the constraint, raising the initial
gas temperature results in lowering the stack temperature.
Single-Pressure Pinch-Constrained HRB Design - Effect of Tg1 on T vs. %Q Profile
(steam @ 50 bar/430 C - 725 psia/806 F)
1292 F
700
Raising Tg1 depresses Tstk
Temperature, C
600
1112
500
932
400
752
300
572
200
392
100
212 F
0
0
10
20
30
40
50
60
70
80
90
∆Teffective
100
Heat Transfer, Q %
Figure 7. Effect of Raising Inlet Gas Temperature on Pinch-Constrained T-Q% Diagrams
The x-axis Fig. 7 is the heat transfer as a percentage of the total heat recovered. The percentages of the total
heat required to heat water in the economiser, then evaporate it, then superheat it to the final temperature are
all fixed, regardless of the inlet gas temperature, since they depend only on steam conditions and properties.
Raising the inlet gas temperature with a fixed pinch results in pivoting the line representing the gas
temperature profile around the pinch point on the T-Q% diagram. Thus, to the extent that the gas
temperature profile can be approximated by a straight line, elevating the hot gas temperature by δTg,1 results
in depressing the stack temperature by δTstack, approximately as follows
δTstack
Heat _ Re quired _ Below _ Pinch _( Economiser )
(1 − φ )
≅
≅
δTg ,1 Heat _ Re quired _ Above _ Pinch _( Evaporator + Superheater )
φ
9-10
(13)
 M. A. Elmasri, 1990-2002
Chapter 9: Single-Pressure HRB/CC Thermodynamics
In the example illustrated above, the economiser needs about 1/3rd of the total heat and the evaporator and
superheater need about 2/3rds of the total heat. Thus, raising the initial gas temperature by 100° depresses
the stack temperature by about 50°.
With pinch-constrained heat recovery, the amount of heat recoverable is proportional to an Effective
Temperature Difference, the difference between the gas inlet temperature and the gas temperature at the
pinch, as illustrated in Fig. 7:
∆Teffective = Tg ,1 − Tg , p = Tg ,1 − (Tsat + ∆T pinch , min )
........... (14)
Since typical exhaust from a heavy duty GT is at about 550 °C (1022 °F), and practical steam saturation
temperatures are around 250 °C (482 °F), the effective temperature difference would be around 280 °C
(about 500 °F). Thus, a 5.6 °C (10 °F) increase in exhaust temperature should result in an increase in heat
recovery (hence in steam generation) of about 5.6/280, i.e. about 2%. For aeroderivative GT’s, the typical
exhaust is at about 450 °C (842 °F), so the effective temperature difference is in the range of 180 °C (about
325 °F), so that a 5.6 °C (10 °F) increase in exhaust temperature should result in an increase in heat recovery
(hence in steam generation) of about 5.6/180, i.e. about 3%.
Naturally, with hotter inlet gas, more heat is recovered, on two counts: (a) the gas is hotter to begin with;
and (b) the stack becomes cooler. This may be seen in Fig. 8, which shows the same diagrams of Fig. 7, but
with the x-axis representing total heat recovery in kW for an exhaust gas flow rate of 100 kg/s.
Single-Pressure Pinch-Constrained HRB Design - Effect of Tg1 on T vs. Absolute Q Profile
(steam @ 50 bar/430 C - 725 psia/806 F)
700
1292 F
Temperature, C
Raising Tg1 depresses Tstk
600
1112
500
932
400
752
300
572
200
392
100
212 F
0
0
10,000
20,000
30,000
40,000
50,000
60,000
Heat Transfer, kW (100 kg/s gas flow)
Figure 8. Effect of Raising Inlet Gas Temperature on Pinch-Constrained T-Q Diagrams
9-11
 M. A. Elmasri, 1990-2002
Chapter 9: Single-Pressure HRB/CC Thermodynamics
3.b. Effect of Inlet Gas Temperature on Stack-Constrained Heat Recovery:
Figure 9 illustrates three T-Q% diagrams, all for generating steam at the same conditions of Fig. 7 [50
bar/430 °C (725 psia/806 °F)], but with flue gases at three different inlet temperatures, 750, 850 and 950 °C
(1382, 1562, and 1742 °F). It is assumed that the minimum stack temperature is 90 °C (194 °F), so that all
three cases are stack limited, since the inlet gas is above the transition temperature at the selected conditions.
Fig. 9 shows that when a single-pressure boiler is designed to produce steam at given conditions from
feedwater at a given condition, and with a given stack temperature as the constraint, raising the initial gas
temperature results in a higher actual pinch, exceeding the minimum pinch constraint.
Single-Pressure Stack-Constrained HRB Design - Effect of Tg1 on T vs. %Q Profile
(steam @ 50 bar/430 C - 725 psia/806 F)
1000
1832 F
900
1652
Raising Tg1 increases the pinch
Temperature, C
800
1472
700
1292
600
1112
500
932
400
752
300
572
200
392
100
212
0
0
10
20
30
40
50
60
70
80
90
∆Teffective
32 F
100
Heat Transfer, Q %
Figure 9. Effect of Raising Inlet Gas Temperature on Stack-Constrained T-Q% Diagrams
Raising the inlet gas temperature with a fixed stack results in pivoting the line representing the gas
temperature profile around the stack temperature on the T-Q% diagram. Thus, to the extent that the gas
temperature profile can be approximated by a straight line, elevating the hot gas temperature by δTg,1 results
in increasing the pinch temperature difference by δ(∆Tpinch), approximately as follows
δ (∆T pinch ) Heat _ Re quired _ Below _ Pinch _( Economiser )
≅
≅ (1 − φ )
δTg ,1
Total _ Heat _ Re quired _
(15)
In the example illustrated above, the economiser needs about 1/3rd of the total heat, so raising the initial gas
temperature by 100° increases the pinch by about 33°.
9-12
 M. A. Elmasri, 1990-2002
Chapter 9: Single-Pressure HRB/CC Thermodynamics
Stack-constrained heat recovery is most likely to be encountered in boilers designed with supplementary
firing. The effective temperature difference is from hot gas inlet all the way to the stack, as shown in Fig. 9.
If the hot gas is at 750 °C (1382 °F) and the stack at 90 °C (194 °F), the effective difference is about 660 °C
(about 1200 °F). Thus, each 5.6 °C (10 °F) increase in hot flue gas temperature should result in an increase
in heat recovery (hence in steam generation) of about 5.6/660, i.e. about 0.8%.
With hotter inlet gas, more heat is recovered and more steam is generated, but on a single count only, viz
that the gas is hotter to begin with. No additional benefit accrues from reducing the stack temperature, as in
the case of pinch-constrained heat recovery, described in §3.a.
Figure 10, which shows the same diagrams of Fig. 9, but with the x-axis representing total heat recovery in
kW for an exhaust gas flow rate of 100 kg/s.
Single-Pressure Stack-Constrained HRB Design - Effect of Tg1 on T vs. Absolute Q Profile
(steam @ 50 bar/430 C - 725 psia/806 F)
1000
1832 F
900
1652
Raising Tg1 increases the pinch
Temperature, C
800
1472
700
1292
600
1112
500
932
400
752
300
572
200
392
100
212
0
0
10,000
20,000
30,000
40,000
50,000
60,000
70,000
80,000
32 F
90,000 100,000
Heat Transfer, kW (100 kg/s gas flow)
Figure 10. Effect of Raising Inlet Gas Temperature on Stack-Constrained T-Q Diagrams
3.c. Effect of Steam Pressure:
Higher steam pressures mean higher saturation temperatures, which diminish the effective temperature
difference if heat recovery is pinch-constrained. Fig. 11 illustrates this fact. If all other parameters are held
constant, designing for a higher pressures rotates the gas temperature profile line upwards on the T-Q%
diagram, around a pivot point at the inlet gas temperature, as seen in Fig. 11 (top).
If the heat recovery is stack-constrained, then higher pressures have no effect on recoverable heat, as shown
by Fig. 12.
9-13
 M. A. Elmasri, 1990-2002
Chapter 9: Single-Pressure HRB/CC Thermodynamics
Single-Pressure Pinch-Constrained HRB Design - Effect of Psteam
600
1112 F
Temperature, C
Raising Psteam elevates Tstk
500
932
400
752
100 bar (1450 psia)
300
572
50 bar (725 psia)
25 bar (363 psia)
200
39
100
212 F
0
0
10
20
30
40
50
60
70
80
90
100
Heat Transfer, Q%
Temperature, C
Single-Pressure Stack-Constrained HRB Design - Effect of Psteam
600
1112 F
500
932
400
752
300
572
200
392
100
212 F
0
0
10,000
20,000
30,000
40,000
50,000
Heat Transfer, kW (100 kg/s gas flow)
Figures 11a & 11b. Effect of Raising Steam Pressure on Pinch-Constrained T-Q% Diagrams (top) and T-Q Diagrams (bottom)
9-14
 M. A. Elmasri, 1990-2002
Chapter 9: Single-Pressure HRB/CC Thermodynamics
Single-Pressure Stack-Constrained HRB Design - Effect of Psteam
Temperature, C
1000
1832 F
900
1652
800
1472
700
1292
600
1112
500
932
400
752
300
572
200
392
100
212
0
0
10
20
30
40
50
60
70
80
90
32 F
100
Heat Transfer, Q%
Single-Pressure Stack-Constrained HRB Design - Effect of Psteam
Temperature, C
1000
1832 F
900
1652
800
1472
700
1292
600
1112
500
932
400
752
300
572
200
392
100
212
0
0
10,000
20,000
30,000
40,000
50,000
60,000
70,000
80,000
32 F
90,000
Heat Transfer, kW (100 kg/s gas flow)
Figures 12a & 12b. Effect of Raising Steam Pressure on Stack-Constrained T-Q% Diagrams (top) and T-Q Diagrams (bottom)
9-15
 M. A. Elmasri, 1990-2002
Chapter 9: Single-Pressure HRB/CC Thermodynamics
3.d. Typical Thermodynamic Trends:
Figures 13 through 15 illustrate typical results for HRB thermodynamic design, as a function of inlet gas
temperature and selected steam pressure. The assumptions used to produce these figures are typical of heat
recovery boilers used for combined cycles, and are summarised and discussed below.
Pinch: The minimum pinch-point temperature difference is assumed at 15 °C (27 °F).
Stack Temperature: The minimum stack temperature is assumed at 90 °C (194 °F).
Approach Subcooling: Ideally, the water supplied from the economiser to the evaporator should be at
saturation temperature. To avoid the possibility of steaming in the economisers, most HRB manufacturers
design the economisers to deliver subcooled water. This "approach subcooling" is usually 3-15 °C F (5-27
°F), with the lower end being used in situations where active off-design controls are applied to suppress
steaming. The subcooling is thermodynamically undesirable in pinch-constrained designs, since it shifts a
little of the heat required from below to above the pinch, reducing steam production capability. An
approach subcooling of 5° C (9 °F) is used in the present calculations.
Minor Losses: Heat loss from the HRB through the insulated duct walls is usually 0.25-1.0 % of heat
recovered, depending on HRB size and extent of insulation. Large boilers have smaller casing surface area
per unit of gas flow and therefore smaller losses. Some gas leakage occurs, mostly by-passing finned tube
sections within the boiler and emerging from the stack. It is generally conservative to assume a 1% heat loss
from a well-constructed HRB. This is assumed in the present calculations.
Blowdown: Another loss from the boiler is blowdown water. Whether this is intermittent or continuous and
the amount depends on water quality and extent of makeup. In the results shown, a 1% continuous
blowdown is assumed. Most modern pure-power combined cycles (no cogeneration steam) will have high
quality water and operate with lower blowdown rates.
Feedwater Temperature: Assumed at 33 °C (91 °F) – consistent with a condenser @ 50 mb (0.725 psia).
Steam Delivery Temperature: Assumed consistent with combined cycle practice, i.e. to obtain the highest
practical steam temperature consistent with the hot flue gas temperature and with material limitations for
piping and steam turbine. Thus, it is assumed to be 540 °C (1004 °F) if the flue gas temperature is above
560 °C (1040 °F), and assumed to be 20 °C (36 °F) below the flue gas temperature, if the flue gas is below
560 °C (1040 °F).
Frequently, towards the cool end of the HRB, a low pressure evaporator generates steam to heat a deaerator
in a closed loop. This results in a jog in the temperature profile but does not alter the thermodynamics
discussed here in any significant way.
Figures 13a & 13b illustrate the thermodynamic fundamentals described above. With hot flue gas
temperatures, above the transition temperature of roughly 700 °C (1292 °F), the heat recovery is stackconstrained (the stack temperature is at its minimum value) and hotter gas results in designs with a pinch
that exceeds the minimum. When the flue gas is below the transition temperature, the heat recovery is
pinch-constrained (pinch is at its minimum value), and cooler gas results in a hotter stack. In the pinchconstrained range, higher steam pressures reduce heat recovery and result in higher stack temperatures.
9-16
 M. A. Elmasri, 1990-2002
Chapter 9: Single-Pressure HRB/CC Thermodynamics
Single-pressure Heat Recovery: Stack Temperature
842 F
932
1022
1112
1202
1382
1292
1472
1562 F
Stack Temperature (C)
240
464 F
220
25 bar (363 psia)
200
100 bar (1450 psia)
428
50 bar (725 psia)
392
180
356
160
320 F
140
284
120
248
100
212
80
450
500
550
600
650
700
750
800
850
Main Assumptions:
Min. pinch = 15 C (27 F)
Min. stack = 90 C (194 F)
Appr. Subcooling=5 C (9 F)
Tfw=33 C (91 F)
Tstm=540 C but <=Tg-20 C
(1004 F but <=Tg-36 F)
176 F
Exhaust Gas Temperature (C)
Single-pressure Heat Recovery: Pinch
842 F
932
Pinch Temperature Difference (C)
80
1022
1112
1202
1382
1292
1472
1562 F
144 F
25 bar (363 psia)
70
126
50 bar (725 psia)
100 bar (1450 psia)
60
108
50
90
40
72 F
30
54
20
36
10
450
500
550
600
650
700
750
800
850
Main Assumptions:
Min. pinch = 15 C (27 F)
Min. stack = 90 C (194 F)
Appr. Subcooling=5 C (9 F)
Tfw=33 C (91 F)
Tstm=540 C but <=Tg-20 C
(1004 F but <=Tg-36 F)
18 F
Exhaust Gas Temperature (C)
Figures 13a & 13b. Stack Temperature (13a-top) & Pinch Temperature Difference (13b-bottom) as Functions of Flue Gas Inlet
Temperature at Different Steam Pressures for Typical HRB Design Parameters
9-17
 M. A. Elmasri, 1990-2002
Chapter 9: Single-Pressure HRB/CC Thermodynamics
Figure 14 shows the typical steam generation rate as a function of flue gas inlet temperature for three values
of steam delivery pressure.
In the pinch-constrained regime, heat recovery is reduced with higher steam pressures. With the hot exhaust
of a heavy duty gas turbine, steam generation is about 14% of exhaust gas flow and weakly dependent on
pressure. For a 575 °C (1067 °F) exhaust, steam production declines by only 7% if design pressure is raised
from 25 bar to 100 bar. With the cooler exhaust of an aeroderivative, steam generation is about 9% of
exhaust flow and depends more strongly on pressure. For a 450 °C (842 °F) exhaust, steam production
declines by 23% if design pressure is raised from 25 bar to 100 bar.
In the stack-constrained regime (flue gas temperatures above the transition), the amount of heat recovered
becomes independent of steam pressure. In this case, steam generation rate is slightly higher at high
pressure because the enthalpy of high-pressure steam is lower than that of low-pressure steam when both are
at the same temperature.
Single-pressure Heat Recovery: Steam Generation
842 F
932
Steam Production, % of Gas Flow
26
1022
1112
1202
1292
1382
1472
1562 F
24
22
Main Assumptions:
Min. pinch = 15 C (27 F)
Min. stack = 90 C (194 F)
Appr. Subcooling=5 C (9 F)
Tfw=33 C (91 F)
Tstm=540 C but <=Tg-20 C
(1004 F but <=Tg-36 F)
20
18
16
14
25 bar (363 psia)
12
50 bar (725 psia)
100 bar (1450 psia)
10
8
6
450
500
550
600
650
700
750
800
850
Exhaust Gas Temperature (C)
Figure 14. Steam Generation as a Function of Flue Gas Inlet Temperature at Different Steam Pressures for Typical HRB
Design Parameters
4. SINGLE-PRESSURE BOTTOMING CYCLES
The term “bottoming cycle” is used to describe a steam cycle which receives its energy input by recovering
heat from another source, such as gas turbine exhaust. In a pure power plant, the bottoming cycle includes a
heat recovery boiler, a steam turbine, a condenser, and a feedpump.
Most of the medium and large bottoming cycles will have dual- or triple-pressure boilers, and the large ones
will also have reheat. Single-pressure, non-reheat bottoming cycles are typically used for small or medium
plants, especially those designed with significant supplementary firing.
9-18
 M. A. Elmasri, 1990-2002
Chapter 9: Single-Pressure HRB/CC Thermodynamics
4.a. Typical Specific Output of Single-Pressure Bottoming Cycles:
Single-pressure Combined Cycle: ST Output per Unit Gas Flow
842 F
932
1022
1112
1202
1382
1292
1472
ST Output, kW per kg/s Gas Flow
340
1562 F
154 kW per lb/s
25 bar (363 psia)
300
136
50 bar (725 psia)
100 bar (1450 psia)
260
118
220
100
180
82
140
64
100
45
60
450
500
550
600
650
700
750
800
Main Assumptions:
Min. pinch = 15 C (27 F)
Min. stack = 90 C (194 F)
Appr. Subcooling=5 C (9 F)
Tfw=33 C (91 F)
Tstm=540 C but <=Tg-20 C
(1004 F but <=Tg-36 F)
Pcond=50 mb (0.725 psia)
27 kW per lb/s
850
Exhaust Gas Temperature (C)
Figure 15. Typical Single-Pressure Bottoming Cycle Steam Turbine Output, per Unit Exhaust Gas Flow to the HRB
Single-pressure Combined Cycle: ST Output per Unit Steam Flow
ST Output, kW per kg/s Steam Flow
842 F
932
1022
1112
1202
1292
1382
1472
1562 F
1250
567 kW per lb/s
1200
544
1150
522
1100
499
25 bar (363 psia)
1050
50 bar (725 psia)
Main Assumptions:
Min. pinch = 15 C (27 F)
Min. stack = 90 C (194 F)
Appr. Subcooling=5 C (9 F)
Tfw=33 C (91 F)
Tstm=540 C but <=Tg-20 C
(1004 F but <=Tg-36 F)
Pcond=50 mb (0.725 psia)
476
100 bar (1450 psia)
1000
950
454
450
500
550
600
650
700
750
800
431 kW per lb/s
850
Exhaust Gas Temperature (C)
Figure 16. Typical Single-Pressure Bottoming Cycle Steam Turbine Output, per Unit Steam Flow to the ST
9-19
 M. A. Elmasri, 1990-2002
Chapter 9: Single-Pressure HRB/CC Thermodynamics
Figures 15 & 16 show the bottoming cycle output that would result if the steam generated under the
assumptions described above (§3.d) were fed to a steam turbine exhausting at a 50 mb (0.725 psia)
condenser. The results shown takes into account the pressure and size effects on the expected steam turbine
isentropic efficiency, and are based on a fixed 300 kg/s (660 lb/s) flue gas mass flow rate entering the boiler;
consistent with a medium GT, such as a GE Frame 7E or an Alstom GT11N, or with a 2-on-1 combined
cycle with a pair of large aeroderivatives, such as LM6000’s or RR Trents. Steam turbines represented in
the diagrams have a nominal output of about 40 MW for plants designed without supplementary firing and
up to about 80 MW for designs incorporating heavy supplementary firing.
With exhaust gas temperatures below about 500 °C (932 °F), although the lowest of the three throttle
pressures gives the lowest output per unit of steam flow, as one may expect, it provides the highest output
per unit of exhaust gas flow, due to the high heat recovery efficiency associated with low steam pressures.
4.b. Bottoming Cycle Efficiency Definitions:
The overall efficiency of a bottoming cycle is defined by the electricity produced as a percentage of the heat
available to the bottoming cycle. By convention, the heat available to the cycle’s HRB is taken to be the
exhaust gas available sensible energy. For an unfired HRB, this is the enthalpy of the hot gases minus their
enthalpy if they were cooled to the ambient temperature without condensing their moisture content (a
hypothetical assumption). For a supplementary fired HRB, the Lower Heating Value (LHV) of the fuel
supplied to the duct burner is added to the denominator:
Steam _ Turbine _ Output
Available _ Sensible _ Energy
WST
=
mg C pg , average (Tg ,1 − Tambient ) + mDB _ fuel LHVDB _ fuel
ηoverall =
...........................
(16)
The overall efficiency combines the efficiency with which heat is recovered, which we shall call the “HRB
Efficiency”,
Heat _ Transferred _ to _ Water / Steam
Available _ Sensible _ Energy
Qwater / steam
=
mg C pg , average (Tg ,1 − Tambient ) + mDB _ fuel LHVDB _ fuel
η HRB =
............................. (17)
with the efficiency with which the recovered heat is converted to power, which we shall call the “Steam
Cycle Internal Efficiency”,
ηint ernal =
So that,
WST
Steam _ Turbine _ Output
=
Heat _ Transferred _ to _ Water / Steam Qwater / steam
η overall = η HRB ∗ηint ernal
............... (18)
........................................................
(19)
The internal efficiency of the steam cycle depends only on the conditions within the steam cycle, such as
throttle pressure, throttle temperature and condenser pressure. It does not address the characteristics of the
heat source and its energy distribution with temperature. To the extent that a high temperature heat source
enables a high throttle temperature, it improves the steam cycle’s internal efficiency.
9-20
 M. A. Elmasri, 1990-2002
Chapter 9: Single-Pressure HRB/CC Thermodynamics
Single-pressure Combined Cycle: Steam Cycle Internal Efficiency
842 F
932
1022
1112
1202
1382
1292
1472
1562 F
Steam Cycle Internal Efficiency, %
37
36
35
Main Assumptions:
Min. pinch = 15 C (27 F)
Min. stack = 90 C (194 F)
Appr. Subcooling=5 C (9 F)
Tfw=33 C (91 F)
Tstm=540 C but <=Tg-20 C
(1004 F but <=Tg-36 F)
Pcond=50 mb (0.725 psia)
34
33
32
25 bar (363 psia)
31
50 bar (725 psia)
100 bar (1450 psia)
30
450
500
550
600
650
700
750
800
850
Exhaust Gas Temperature (C)
Figure 17. Typical Single-Pressure Bottoming Cycle Internal Efficiencies
Single-pressure Combined Cycle: Heat Recovery Efficiency
842 F
932
1022
1112
1202
1292
1382
1472
1562 F
95
90
HRB Efficiency, %
85
80
Main Assumptions:
Min. pinch = 15 C (27 F)
Min. stack = 90 C (194 F)
Appr. Subcooling=5 C (9 F)
Tfw=33 C (91 F)
Tstm=540 C but <=Tg-20 C
(1004 F but <=Tg-36 F)
75
70
65
60
25 bar (363 psia)
55
50 bar (725 psia)
50
45
100 bar (1450 psia)
450
500
550
600
650
700
750
800
850
Exhaust Gas Temperature (C)
Figure 18. HRB Efficiency as a Function of Flue Gas Inlet Temperature at Different Steam Pressures for Typical HRB Design
Parameters
9-21
 M. A. Elmasri, 1990-2002
Chapter 9: Single-Pressure HRB/CC Thermodynamics
Figure 17 shows how steam cycle internal efficiencies increase with throttle pressure, within the practical
range, and also increase with throttle temperature. Once the flue gas is hot enough to permit the maximum
practical throttle temperature, its only (weak) influence on cycle internal efficiency derives from the fact that
the design points in Fig. 17 are computed with a fixed exhaust flow rate. Thus, with hotter gas, the steam
turbine is larger in its throttle volumetric flow rate and its rating. This implies a machine with a better
generator efficiency and higher isentropic efficiency, as discussed in §4.xxx below, especially with higher
throttle pressures.
Figure 18 shows HRB efficiencies for typical design assumptions. If the flue gas is hot enough to attain the
minimum stack temperature, HRB efficiency is independent of steam pressure, and is very high, in the
neighbourhood of 90% for a 90 °C (194 °F) stack. For unfired exhaust, which is typically below the
transition temperature HRB efficiency depends on steam pressure, and is in the 75% range for exhaust
temperatures normally associated with heavy duty GT’s, and in the 60% range for exhaust temperatures
normally associated with aeroderivative GT’s.
Figure 19 shows how the overall bottoming cycle efficiency depends strongly on the exhaust gas
temperature, since hotter gas improves HRB efficiency, at least within the pinch-controlled design range, as
well as helping to improve the steam cycle internal efficiency.
Single-pressure Combined Cycle: Steam Cycle Overall Efficiency
842 F
932
1022
1112
1202
1382
1292
1472
1562 F
Steam Cycle Overall Efficiency, %
35
30
Main Assumptions:
Min. pinch = 15 C (27 F)
Min. stack = 90 C (194 F)
Appr. Subcooling=5 C (9 F)
Tfw=33 C (91 F)
Tstm=540 C but <=Tg-20 C
(1004 F but <=Tg-36 F)
Pcond=50 mb (0.725 psia)
25
25 bar (363 psia)
50 bar (725 psia)
20
15
100 bar (1450 psia)
450
500
550
600
650
700
750
800
850
Exhaust Gas Temperature (C)
Figure 19. Bottoming Cycle Overall Efficiency as a Function of Flue Gas Inlet Temperature at Different Steam Pressures for
Typical Design Parameters
4.c. Thermodynamically Optimum Throttle Pressure:
Figure 18 shows that with low exhaust gas temperatures, a low throttle pressure, such as 25 bar (363 psia),
provides the best performance. This is because with pinch-constrained heat recovery a low HRB pressure
allows a low stack temperature and hence efficient heat recovery (Fig. 18). The efficiency of the heat
recovery more than compensates for the fact that low pressure steam produces less power per unit of steam
9-22
 M. A. Elmasri, 1990-2002
Chapter 9: Single-Pressure HRB/CC Thermodynamics
mass flow rate, hence low steam cycle internal efficiency (Fig. 17). With higher gas temperatures, on the
other hand, heat recovery can be efficient even with high steam pressures. If the gas is hot enough to
transition into the stack-constrained regime, the efficiency of heat recovery becomes independent of
pressure (Fig. 18). Thus, the thermodynamically optimum pressure increases with gas temperature.
Steam turbine isentropic efficiency is a function of size, primarily steam volumetric flow rate. Small steam
turbines tend to have lower efficiencies due to their shorter blading, which results in higher end-wall and
stage leakage losses. Thus, all other parameters being equal, ST isentropic efficiency is higher for larger
turbines.
Higher throttle pressures give a fundamental thermodynamic advantage, due to the greater enthalpy drop
between the throttle and the condenser pressure. On the other hand, high pressures reduce turbine isentropic
efficiency because (a) they reduce the volumetric flow rate in the HP stages, and (b) they increase steam
wetness in the LP stages, as seen in Figure 20. With wet steam, isentropic efficiency is reduced due to the
moisture droplets inability to follow the steam streamlines, which causes them to strike the turbine blades at
incorrect angles.
Specific Enthalpy, h
Throttle temperature
High throttle pressure
Lower throttle
pressure
P condenser
Higher x (drier steam))
Low x (wet steam)
Specific Entropy , s
Figure 20. A high throttle pressure creates a longer expansion line to the condenser pressure which tends to increase power
output per unit of steam flow, but results in greater wetness in the LP stages
Figure 21 shows how throttle pressure affects bottoming cycle overall efficiency, as a function of system
size. For a steam turbines, in the 40-90 MW class, optimum pressure is about 40 bar (600 psia), whereas for
small turbines in the 15 MW class it is about 30 bar (450 psia).
9-23
 M. A. Elmasri, 1990-2002
Chapter 9: Single-Pressure HRB/CC Thermodynamics
Bottoming Cycle Pressure Optimisation
Bottoming Steam Cycle Overall Efficiency, %
290
580 psia
870 psia
1160
25.4
25.2
25.0
600 kg/s (1320 lb/s) Gas Flow
(90 MW Class ST)
24.8
300 kg/s (660 lb/s) Gas Flow
(45 MW Class ST)
24.6
24.4
100 kg/s (220 lb/s) Gas Flow
(15 MW Class ST)
24.2
24.0
Same assumptions as Fig. 19
Flue Gas @ 550 C (1022 F)
23.8
23.6
23.4
20
40
60
80
Throttle Pressure (bar)
Figure 21. Size Effect on Optimum Throttle Pressure
Single-Pressure Bottoming Cycle Exhaust Energy Distribution
at the Optimum Throttle Pressure (40 bar / 580 psia)
Misc. Losses
1.35%
Stack Loss
25.67%
Condenser
Loss
47.92%
ST Output
25.06%
Figure 22. Typical Exhaust Energy Distribution at the Optimum Throttle Pressure
Figures 22 & 23 help explain the thermodynamic trends as the selection of the design pressure is varied
from the optimum for a bottoming cycle in the 45 MW class. Fig. 23 shows that at the optimum 40 bars, the
stack loss is about 26%, with the HRB recovering about 74% of the heat (Fig. 18). The steam cycle internal
9-24
 M. A. Elmasri, 1990-2002
Chapter 9: Single-Pressure HRB/CC Thermodynamics
efficiency (Fig. 17) is about 34%, so 0.34 x 0.74 ≈ 25% of the heat is converted to power, for an overall
efficiency of about 25% (Fig.19). The portion of the recovered heat which is not converted to power, about
74%-25%=49%, is rejected at the condenser.
Selecting a pressure lower than optimum allows a lower stack temperature and greater heat recovery.
However, since the steam is at lower pressure, the expansion line is shorter and less of its energy is
converted to power, but more is rejected to the condenser. Conversely, selecting a pressure higher than
optimum raises stack temperature, increasing energy loss at the stack, but increases the work per unit mass
of steam, due to the longer expansion line, and reduces condenser heat rejection. Fig. 23 shows the
variations in the energy balance relative to the base case with the optimum throttle pressure.
Changes in Exhaust Energy Distribution with Throttle Pressure for 1-P Bottoming Cycle,
Relative to the Optimim Throttle Pressure of 40 bar (580 psia)
Change in Energy Distribution, as % of Total Exhaust Gas Energy
5
4
3
2
1
Stack Loss
ST Output
0
Condenser Loss
Misc Losses
-1
-2
-3
-4
20 bar
290 psia
30 bar
435 psia
40 bar
580 psia
50 bar
725 psia
60 bar
870 psia
70 bar
1015 psia
80 bar
1160 psia
-5
ST Throttle Pressure
Figure 23. Changes in Bottoming Cycle Energy Balance with Throttle Pressure Selection
The economically optimum pressure would tend to be higher, not lower, than the thermodynamic optimum.
Higher pressure means less heat recovery and less heat rejection, with a greater stack loss. Rejecting heat at
the stack is cheaper than recovering it in the HRB then rejecting it at the condenser.
4.d. Other Practical Constraints on Throttle Pressure:
In addition to the drop in stage efficiency with steam wetness, excessive moisture can physically damage the
turbine blades by erosion. In most practical designs, the maximum practical wetness is 10-12% (minimum
steam quality 88-90%). Steam turbines designed to operate with wetter steam, such as in nuclear cycles,
9-25
 M. A. Elmasri, 1990-2002
Chapter 9: Single-Pressure HRB/CC Thermodynamics
would be designed with lower flow velocities, which increase number of stages and hence cost, and would
have mechanisms for moisture extraction from various bleeds along the expansion path. Most steam
turbines for combined cycle use are not equipped with moisture extraction, so the combination of design
throttle pressure and condenser pressure should be selected to keep moisture at the end of the expansion line
below 12%. This places an upper bound on throttle pressure which depends on throttle temperature,
condenser pressure, and turbine efficiency.
Another practical constraint arises from the need for a minimum volumetric flow rate of steam consistent
with the turbine’s rotational speed. To achieve sound aerodynamics and a reasonable stage pressure ratio,
the blade peripheral speed (u=πDN/60 where D is diameter and N RPM) for the first turbine stage should
ideally be in the neighborhood of 160 m/s (520 ft/s), with the acceptable range being 110-210 m/s (360-690
ft/s). For a particular rotational speed, this fixes the desirable pitch diameter and its acceptable range. The
diameter, together with a reasonable blade height in proportion to the diameter and an appropriate axial flow
velocity, all determine the desirable range of steam volumetric flow rate for any given rotational speed.
These considerations result in the following throttle pressures for synchronous steam turbines (3000 or 3600
RPM), dependent on size:
Turbine MW Class
20
50
100
350
Throttle Pressure Range
bar
psia
25-60
360-900
40-90
600-1300
50-120
725-1750
100-250
1450-3600
5. THERMODYNAMICS OF SUPPLEMENTARY FIRING
5.a. Duct Burner Fuel Utilisation Efficiency:
Fig. 18 shows how the HRB efficiency varies with gas temperature and steam pressure. For the 50 bar (725
psia) steam, about 80% of the energy can be recovered from a 600 °C (1112 ° F) gas, compared to only
about 60% of the energy from a 450 °C (842 °F) gas. This shows the importance of high exhaust
temperature for gas turbines to be used in high-grade heat recovery applications. Not only is there more heat
to recover, but a bigger percentage is recoverable if the heat is of higher grade. This begs the question
whether fuel can be utilised in a duct burner at an efficiency exceeding 100%, since raising the gas
temperature not only raises its energy (in proportion to the DB fuel consumed), but also raises the
percentage of the total gas energy (the original energy and that added by the duct burner) that can be
recovered.
We define duct burner fuel utilisation efficiency as
η DBFU =
Qre cov ered , w / sup . firing − Qre cov ered ,unfired
Total _ Fuel _ Input _ to _ DB _( LHV )
............................................. (20)
and incremental duct burner fuel utilisation efficiency as
η DBFU ,i =
Incremental _ Heat _ Re cov ered
Incremental _ Fuel _ Input _ to _ DB _( LHV )
9-26
................................ (21)
 M. A. Elmasri, 1990-2002
Chapter 9: Single-Pressure HRB/CC Thermodynamics
These are plotted in Fig. 24 for three different HRB pressures, against the temperature rise by duct firing of
exhaust gas initially at 550 °C (1022 °F). As long as the DB exit temperature is below the transition
temperature (i.e. as long as the heat recovery is pinch-constrained), the incremental and the total DBFU
efficiencies are the same, and well above 100%. This is because with pinch-constrained heat recovery,
increasing hot gas temperature by δTg,1 allows a decrease in stack temperature of δTstack = [ϕ/(1-ϕ)] δTg,1
according to Equation 13. Since ϕ (the proportion of the total heat that goes to the economiser) increases
with pressure, so does the DBFU efficiency.
Duct Burner Fuel Utilisation Efficiency - Design Points with Unfired Gas @ 550 C (1022 F)
1022
160
1112
1202
1292 F
1382
1472
1562
155
150
DBFU Efficiency, %
145
140
25 bar (total DBFU)
135
25 bar (incremental)
130
50 bar (total DBFU)
125
50 bar (incremental)
120
100 bar (total DBFU)
100 bar (incremental)
115
110
105
100
95
550
600
650
700
750
800
850
Gas Temperature Entering HRB, after DB (C)
Figure 24. Typical Duct Burner Fuel Utilisation Efficiencies
Once the extent of supplementary firing raises the gas to the transition temperature, the incremental DBFU
efficiency drops to just under 100%. After transition to the stack-constrained regime, increasing Tg,1 can no
longer depress Tstack so the DB incremental heat input is all converted to additional steam, except for the
HRB minor losses (assumed @1%) and the effect of the slight increase in flue gas flow rate to the stack.
The total DBFU efficiency begins also to drop, since it is the average of all the increments, but still remains
well above 100% for the range of interest.
5.b. Effect of Supplementary Firing on Single-Pressure Combined Cycle Efficiency:
The efficiency of converting duct burner fuel to power is the product of the efficiency of converting duct
burner fuel to steam (DBFU efficiency) and the efficiency of converting steam to power (the steam cycle
internal efficiency), i.e.,
η DB _ fuel → Power = η DBFU ∗ηi
..................................................... (22)
If this exceeds the base efficiency of an unfired combined cycle, then duct firing will improve cycle
efficiency. If this is below the base efficiency of an unfired combined cycle, then duct firing will lower
cycle efficiency.
9-27
 M. A. Elmasri, 1990-2002
Chapter 9: Single-Pressure HRB/CC Thermodynamics
In most cases, the efficiency given by Equation (22) is very close to that of a single-pressure, unfired
combined cycle, so the effect of designing a single-pressure CC with supplementary firing is close to neutral
on efficiency. For example, consider the Frame 6B single-pressure CC shown below. Net efficiency is
47.33%.
Gross Power
Net Power
Aux. & Losses
Gross Heat Rate
Net Heat Rate
Gross Electric Eff.
Net Electric Eff.
Fuel LHV Input
Fuel HHV Input
Net Process Heat
62024 kW
60811 kW
1213.2 kW
7458 kJ/kWh
7607 kJ/kWh
48.27 %
47.33 %
128490 kWth
142574 kWth
0 kWth
Ambient
1.013 P
15 T
60%RH
50 p
525 T
18.3 M
Stop Valve
20675 kW
0.05 p
32.9 T
18.12 M
1.013 p
162.2 T
145.2 M
to HRSG
HP
CH4
2.567 M
128490 kWth LHV
HPB
51.5 p
265.8 T
18.3 M
1.003 p
15 T
142.6 M
469.8 T
280.8 T
1.038 p
547.2 T
145.2 M
33.1 T
18.49 M
p [bar] T [°C] M [kg/s], Steam Properties: Thermoflow - STQUIK
0 05-08-2002 19:51:42 file=C:\SEMINAR\Sem_5-02\1-p\GTP_4-...ase_1PFr6CC.gtp
GE 6581B
@100%load
41349 kW
GT PRO 10.3.2 Maher Elmasri
If designed with supplementary firing to 650 °C, the net efficiency is 47.24%, as shown below.
Design with Supplementary Firing to 650 C
Gross Power
70803 kW
Net Power
69347 kW
Aux. & Losses
1456.1 kW
Gross Heat Rate
7463 kJ/kWh
Net Heat Rate
7620 kJ/kWh
Gross Electric Eff.
48.24 %
Net Electric Eff.
47.24 %
Fuel LHV Input
146782 kWth
Fuel HHV Input
162871 kWth
Net Process Heat
0 kWth
Ambient
1.013 P
15 T
60%RH
50 p
525 T
25.88 M
Stop Valve
29454 kW
0.05 p
32.9 T
25.62 M
1.013 p
113.6 T
145.6 M
to HRSG
HP
CH4
2.567 M
128490 kWth LHV
HPB
51.5 p
265.8 T
25.88 M
1.003 p
15 T
142.6 M
543.6 T
280.8 T
1.038 p
547.2 T
145.2 M
33.1 T
26.14 M
CH4
0.3655 M
18292 kWth LHV
p [bar] T [°C] M [kg/s], Steam Properties: Thermoflow - STQUIK
0 05-08-2002 19:54:13 file=C:\SEMINAR\Sem_5-02\1-p\GTP_4-02\Fired_1PFr6CC.gtp
GE 6581B
@100%load
41349 kW
GT PRO 10.3.2 Maher Elmasri
9-28
 M. A. Elmasri, 1990-2002
Chapter 9: Single-Pressure HRB/CC Thermodynamics
The DBFU efficiency, from Fig. 24, is 138% and steam cycle internal efficiency, from Fig. is 34%. Thus,
η DB _ fuel → Power = η DBFU ∗ηi = 1.38 ∗ 0.34 = 0.4692 = 46.92%
which is just below the base case unfired design, hence the very small, almost negligible, reduction in
efficiency.
If the same exercise given above were repeated with an older, less efficient GT model, such as a GE6541
(instead of a 6581), supplementary firing would be seen to very slightly increase CC efficiency. This is
especially true if the supplementary firing were to enable a throttle temperature higher than would be
otherwise attainable, thus increasing the steam cycle’s internal efficiency relative to that in an unfired
design.
6. HRB SURFACE AREA CALCULATION
6.a. Principles:
In design calculations, such as those described in this chapter, the designer selects the key thermodynamic
variables for the cycle, such as throttle pressure and temperature and condenser pressure; as well as the key
thermodynamic design assumptions, such as the minimum pinch-point temperature difference and the
minimum stack temperature. These design criteria and assumptions permit the heat and mass balance to be
solved, and results similar to those shown above to be obtained. Once the flows, pressure and temperatures
at each point in the cycle have been found, the next step is to calculate the equipment sizes that would be
necessary to implement the solution.
For any heat exchanger in the HRB, the heat transfer rate Q, surface area A, and a suitably-averaged
temperature difference ∆T are related through the overall heat transfer coefficient U, as follows:
Q = UA∆T ........... (23)
∴ A = Q /(U ∆T )
(23a)
To the extent that the temperature profiles
within the heat exchanger can be
approximated by straight lines, as shown
in Fig. 25, the average temperature
difference can be approximated by the
Logarithmic
Mean
Temperature
Difference (LMTD), which is defined in
terms of the temperature differences at the
two ends of the heat exchanger as follows:
∆T ≅ LMTD =
∆T1
∆T2
∆T1 − ∆T2
... (24)
n(∆T1 / ∆T2 )
If either ∆T1 or ∆T2 is zero, the LMTD is
zero. If ∆T1 = ∆T2, the LMTD is equal to
them. If ∆T1 and ∆T2 are unequal and
Figure 25
9-29
 M. A. Elmasri, 1990-2002
Chapter 9: Single-Pressure HRB/CC Thermodynamics
non-zero, which is likely in practice, the LMTD is smaller than the arithmetic mean of ½(∆T1+∆T2) and
becomes increasingly so as the disparity between If ∆T1 and ∆T2 increases.
The overall heat transfer coefficient (U) depends on the gas and water properties, their flow velocities, the
size, geometry and materials of the boiler tubes and fins, fouling factors, etc. Roughly twenty design
parameters need to be selected for each heat exchanger to calculate its surface area. In addition, it is
common to design the boiler subject to a gas-side pressure drop criterion, which influences the final surface
area needed along with the heat transfer duty and temperatures.
6.b. Typical Results:
Figures 26 through 28 shows specific HRB surface areas for a design gas-side pressure drop of 20 mb (8"
water). When a boiler is designed with higher gas temperatures (and correspondingly lower stack
temperatures in the pinch-constrained range), it requires more heat transfer area due to the increased heat
transfer (Q in Equation 23). If the design point hot gas temperature is in the vicinity of the transition value,
where both pinch and stack are at or near their minima, the required surface area is very large, mostly
because of the need for a very large economiser, as shown in Fig. 27. This is because the gas and water
temperature profiles over the economiser approach close, parallel lines, with a small ∆T prevailing for a
large portion of the heat transfer (as seen in Fig. 7). With higher gas temperatures, in the stack-constrained
range, the pinch becomes larger than its minimum value, and the required surface area is reduced. If a boiler
will be sized to operate with gas in the vicinity of the transition temperature, it is more cost effective to
design it with a larger pinch.
Single-pressure Heat Recovery: Required Boiler Area per Unit Gas Flow
842 F
932
Boiler Area, sq.m per kg/s Gas Flow
360
1022
1112
1202
1382
1292
1472
1562 F
1758 sq.ft per lb/s
25 bar (363 psia)
50 bar (725 psia)
300
1465
100 bar (1450 psia)
240
1172
180
879
120
586
60
450
500
550
600
650
700
750
800
Main Assumptions:
Min. pinch = 15 C (27 F)
Min. stack = 90 C (194 F)
Appr. Subcooling=5 C (9 F)
Tfw=33 C (91 F)
Tstm=540 C but <=Tg-20 C
(1004 F but <=Tg-36 F)
DPgas=20 mb (8" H2O)
293 sq.ft per lb/s
850
Exhaust Gas Temperature (C)
Figure 26. Typical Single-Pressure HRB Required Surface Areas per unit of Gas Flow
Higher pressures require larger HRB areas, due to the decreasing LMTD’s for the evaporator and
superheater, which outweigh the increasing LMTD for the economiser, except at low gas temperatures,
9-30
 M. A. Elmasri, 1990-2002
Chapter 9: Single-Pressure HRB/CC Thermodynamics
where the reduction in heat transfer with increasing pressure reduces the overall area required. Please
review Figs. 11 and 12 to clarify these trends.
Single-pressure Heat Recovery: Required Boiler Area per Unit Gas Flow
842 F
932
1022
1112
1202
1382
1292
1472
1562 F
Boiler Area, sq.m per kg/s Gas Flow
300
1465 sq.ft per lb/s
Total Area
Economiser
Evaporator
Superheater
240
1172
180
879
120
586
60
293
0
450
500
550
600
650
700
750
800
850
Main Assumptions:
50 bar (725 psia)
Min. pinch = 15 C (27 F)
Min. stack = 90 C (194 F)
Appr. Subcooling=5 C (9 F)
Tfw=33 C (91 F)
Tstm=540 C but <=Tg-20 C
(1004 F but <=Tg-36 F)
DPgas=20 mb (8" H2O)
0 sq.ft per lb/s
Exhaust Gas Temperature (C)
Figure 27. Typical Single-Pressure HRB Required Surface Area Breakdown for Economiser, Evaporator, and Superheater
9-31
 M. A. Elmasri, 1990-2002
Chapter 9: Single-Pressure HRB/CC Thermodynamics
Single-pressure Heat Recovery: Required Boiler Area per Unit Steam Production
842 F
Boiler Area, sq.m per kg/s Steam Generated
1800
932
1022
1112
1202
1292
1382
1472
1562 F
8789 sq.ft per lb/s
1600
7813
1400
6836
1200
5859
1000
4883
25 bar (363 psia)
50 bar (725 psia)
800
600
3906
100 bar (1450 psia)
450
500
550
600
Main Assumptions:
Min. pinch = 15 C (27 F)
Min. stack = 90 C (194 F)
Appr. Subcooling=5 C (9 F)
Tfw=33 C (91 F)
Tstm=540 C but <=Tg-20 C
(1004 F but <=Tg-36 F)
DPgas=20 mb (8" H2O)
650
700
750
800
850
2930 sq.ft per lb/s
Exhaust Gas Temperature (C)
Figure 28. Typical Single-Pressure HRB Required Surface Area per unit of Steam Generation
6.c. Designing a Boiler to Operate at Off-Design Across a Range of Gas Temperatures:
Boilers with supplementary firing have to work satisfactorily across a range of gas temperatures. For
example, with a GT exhaust at 500 °C (932 °F), a supplementary fired boiler may be equipped with a duct
burner capable of raising the gas to 800 °C (1472 °F), doubling the steam production, as seen in Fig. 14.
Figure 27 [for 50 bar (725 psia) steam] shows the difficulty encountered in sizing the various components of
such a boiler. If the design point were the unfired case, the total surface area needed (per kg/s of gas flow)
is about 135 m2 whereas it would be about 190 m2 if the design-point were the fully-fired case. The
discrepancies for the economiser and superheater are much greater. Designed at the unfired case, the
economiser would be sized at about 47 m2 per kg/s of gas, less than half its size if it were designed at the
fully-fired case 133 m2 per kg/s of gas. The opposite holds for the superheater. If it were designed at the
unfired case, it would be sized at roughly 23 m2 per kg/s of gas, about triple its size of 7.6 m2 per kg/s of gas
if it were based upon the fully-fired case. Also, a superheater designed for high temperature gas would be
made out of high alloy steel, such as T91 or 409 stainless, but one designed for moderate temperature would
be made of less expensive alloy, such as T22. Furthermore, a superheater designed for high temperature gas
should have lower fin density, shorter fins (or even bare tubes) compared to one designed for a moderate
temperature.
The solution to this problem is to select materials and fin geometry based on the maximum level of
supplementary firing, then to calculate the necessary surface areas at some intermediate nominal case. After
testing the performance at the extremes of the operating range, the initial design may be modified if
necessary. This process is illustrated by Figures 29a-29c.
9-32
 M. A. Elmasri, 1990-2002
Chapter 9: Single-Pressure HRB/CC Thermodynamics
Figure 29a shows the temperature profile
at a sizing point with gas at 650 °C (1202
°F). The economiser is designed for 5 °C
(9 °F) exit subcooling and the superheater
for 550 °C (1022 °F) steam. The pinch is
assumed at 15 °C (27 °F). The boiler is
sized at that point, then its components are
fixed and their performance at the
extremes of the range, 800 °C (1472 °F)
and 500 °C (932 °F) is tested.
Figure 29a
When maximum firing is used, the economiser will appear undersized and will deliver cooler water to the
evaporator than at the nominal case, and the superheater will appear oversized and deliver hotter steam than
nominal, which can then be desuperheated with a water spray.
9-33
 M. A. Elmasri, 1990-2002
Chapter 9: Single-Pressure HRB/CC Thermodynamics
In our example, the T-Q diagram for the
off-design case with ‘maximum firing’ is
shown in Fig. 29b. The economiser exit
temperature has dropped by 17 °C (31 °F)
from the design point, its exit subcooling
increasing from 5 °C (9 °F) to 22 °C (40
°F). Superheater exit temperature has
risen by 74 °C (133 ° F) to 624 °C (1155
°F). A water spray of about 5.5% needs to
be used to desuperheat the steam to the
nominal temperature. The pinch-point
temperature difference has increased from
the nominal 15 °C (27 °F) to 20.5 °C (37
°F). The stack temperature has dropped
from the nominal 118 °C (244 °F) to 95°
C (203 °F). (Note that the drop in stack
temperature when hotter gas is imposed
upon a fixed boiler is less than the drop
that would occur if the boiler were redesigned with the hotter gas.)
Figure 29b
When operating unfired, the economiser will appear oversized and will deliver warmer water than the
nominal case (or may it may steam), and the superheater will appear undersized and its exit temperature will
be below nominal. If the latter is unacceptable, the superheater can be made large enough to give the correct
steam temperature at the unfired condition, but will then require a larger amount of desuperheating water at
the fully fired condition.
In our example, the T-Q diagram for the
off-design ‘unfired’ case is shown in Fig.
29c. The economiser exit subcooling
has vanished. In fact, the economiser is
steaming, with an exit quality of 2.3%.
If this is unacceptable, the economiser
would have to be reduced in size, but
this would reduce performance at the
nominal operating condition.
The
superheater exit temperature has dropped
to 461 °C (864 °F).
If this is
unacceptably low, the superheater would
need to be enlarged, but that would mean
greater desuperheating flows at higher
gas temperatures, which have a negative
impact on efficiency. At this off-design
condition, the pinch has shrunk from the
nominal 15 °C (27 °F) to 9.6 °C (17 °F),
and the stack temperature has climbed
from the nominal 118 °C (244 °F) to 158
°C (316 °F).
Figure 29c
9-34
 M. A. Elmasri, 1990-2002
Chapter 9: Single-Pressure HRB/CC Thermodynamics
7. EFFECT OF KEY PARAMETERS
7.a. Effect of Pinch Temperature Difference in a Pinch-Constrained Design:
Reducing the pinch by x° has the effect of increasing the effective temperature difference, shown in Figs. 7
& 9 by x°. This increases steam production (and power of a bottoming cycle) by x°/∆Teffective (both being in
the same units °C or °F). Thus, if for typical unfired exhaust we assume that ∆Teffective about 300 °C, a 3 °C
reduction in pinch would increase steam production by about 1%. Thus, with high pressures or low gas
temperatures, the percentage gain in steam production for each degree of pinch reduction increases.
The effect of changing the pinch on the stack temperature can be approximated by
δTstack
δ (∆T pinch )
≅
1
Total _ Heat _ Re quired
≅
Heat _ Re quired _ Above _ Pinch _( Evaporator + Superheater ) φ
(25)
so that a 3° reduction in the pinch would correspond to a 4°-5° reduction at the stack.
Reducing the pinch has an exponential effect on increasing the required boiler surface area, which becomes
infinitely large as the pinch approaches zero. Reducing the pinch from 15 °C to 5 °C requires a HRB with
about 50% more area but only adds about 3-4% to steam production and CC output. Selecting the pinch for
optimising plant economics is discussed elsewhere.
7.b. Effect of Economiser Approach Subcooling:
In pinch-constrained heat recovery, the approach subcooling reduces steam output because the subcooled
water needs to be heated to the saturation temperature within the evaporator, using flue gas heat above the
pinch, rather than within the economiser, using flue gas heat below the pinch. Approach subcooling raises
the stack temperature, since heat that would have been taken between the pinch and the stack is taken from
above the pinch instead. Thus, neglecting blowdown steam output is reduced in the ratio
∆hsubcooling/(hs-hf)
where hs is the final steam enthalpy and hf the enthalpy of saturated water.
With a subcooling of 5 °C (9 °F), ∆hsubcooling is about 25-30 kJ/kg (10-13 BTU/lb), dependent upon (and
increasing with) pressure. The enthalpy increase above the pinch is typically in the range 1800-2600 kJ/kg
(775-1175 BTU/lb), dependent upon (and decreasing with) pressure, and dependent upon (and increasing
with) steam temperature. Thus, approach subcooling typically reduces steam output by 1-1.5%. This
penalty is avoided by once-through boilers, and by circulating-drum boilers designed to operate
satisfactorily with no approach subcooling or mild economiser steaming.
7.c. Effect of Steam Temperature:
If all other parameters are held constant, increasing the steam temperature will increase the proportion of the
total heat needed above the pinch, thereby reducing steam output and raising stack temperature in the pinchconstrained range. In spite of the reduction in steam flow, power output of a steam turbine will increase,
since the higher enthalpy at the throttle will outweigh the effect of the reduction in steam flow rate. In the
overall bottoming cycle energy balance, the reduction in condenser heat rejection due to the decreased steam
flow will exceed the increase in stack heat loss occasioned by the higher stack temperature.
9-35
 M. A. Elmasri, 1990-2002
Chapter 10: Dual-Pressure HRB/CC Thermodynamics
THERMODYNAMICS OF DUAL-PRESSURE HEAT RECOVERY
BOILERS & COMBINED CYCLES
Content Revised May, 2002
© Maher Elmasri 1990-2002
1. THERMODYNAMICS OF DUAL-PRESSURE HEAT RECOVERY
1.a. Rationale for Dual-Pressure Boilers:
The pinch constraint reduces the effectiveness of single-pressure heat recovery from a moderate temperature
source, such as typical, unfired, GT exhaust. For high steam pressures, this constraint forces a high stack
temperature, which results in low heat recovery efficiency. With lower pressures, heat recovery is efficient,
but the steam is less-useful for producing power.
A logical improvement is to recover heat at high pressure, followed by a second, lower pressure boiler to
scavenge more heat from the gases, rather than let this excess heat go up the stack.
This is illustrated in Figure 1. The High Pressure (HP) is followed by a second boiler, called Intermediate
Pressure (IP). This nomenclature helps to distinguish it from a possible third, Low Pressure (LP) boiler that
may be used for deaeration or in Triple-Pressure systems.
The arrangement shown in Figure 1 uses a combined HP/IP economiser at the cool end of the boiler. Water
is pressurised to the HP level in a single pump, then it is heated in the combined economiser. A part of this
water is let-down to the IPB. This simplifies the design, but wastes a few kW in the pump. There are many
variations to this arrangement at the cooler end of the boiler, each with practical merits as well as
disadvantages, but all with essentially the same thermodynamic results.
Figure 1. Layout of a basic Dual-Pressure Heat Recovery Boiler
10-1
 M. A. Elmasri, 1990-2002
Chapter 10: Dual-Pressure HRB/CC Thermodynamics
Figure 2 shows the temperature
profile of the HRB shown in
Figure 1.
With the assumed
feedwater and steam conditions,
pinch-points at 15 °C (27 °F), and
a hot flue gas temperature of 520
°C (968 °F), the stack temperature
is 116 °C (241 °F), above the 90
°C (194 °F) which we assume
possible for clean gas fuel.
Min ∆Tpinches of
15 C both achieved
Min Tstk of
90 C not
achieved
Figure 2. Example of a temperature profile for a Dual-Pressure boiler with
Pinch-Constrained heat recovery
The IP in a two-pressure system is of secondary importance to the HP. Typically, with unfired exhaust, IP
steam production is about one-quarter that of the HP; the IP steam is also of lower value. Because the IP
recovers the "left-over" heat missed by the HP, changes in any parameter that tend to increase HP heat
recovery, would tend to decrease IP heat recovery and vice versa.
1.b. Pinch & Stack Constraints:
In single-pressure heat recovery we found two constraints that limit the extent to which the gas may be
cooled, and hence the amount of heat which may be recovered. These were the minimum stack temperature
and the minimum temperature difference at the pinch. If the temperature profile allowed the minimum stack
temperature to be attained, the gas energy between the inlet and the stack determined the heat recovery and
steam generation. The corresponding temperature difference at the pinch could then be computed, and
would exceed its minimum value. If the temperature profile did not allow the minimum stack temperature
to be reached, the heat recovery was pinch-constrained. The heat recovery and steam generation were
determined by the gas energy between the inlet and the pinch. The corresponding stack temperature could
then be computed, and would exceed its minimum value. For given feedwater and steam conditions, there
was one inlet gas temperature at which both minimum stack and minimum pinch were simultaneously
satisfied. This was called the transition temperature. If the inlet gas were cooler, heat recovery was pinchconstrained. If the inlet gas were hotter, heat recovery was stack-constrained.
With dual-pressure heat recovery there are three constraints instead of two. Minimum stack, minimum HP
pinch, and minimum IP pinch. There are two unknown flow rates: HP steam generation and IP steam
generation. Thus, there are three regimes described in 1.c through 1.e below.
10-2
 M. A. Elmasri, 1990-2002
Chapter 10: Dual-Pressure HRB/CC Thermodynamics
1.c. The Pinch-Constrained regime:
Both HP and IP pinches are at their minimum values. The two equations for gas energy balance above each
pinch can be solved to yield the two unknown steam/water flow rates. Once these are found, the energy
balance to the stack can be used to find the stack temperature. If this is above the minimum allowed value,
then this is indeed the correct regime and correct solution. The maximum inlet gas temperature for this
regime is the value at which the two pinch-point temperature differences and the stack temperature can all
be at their minimum values simultaneously. We call this the dual-pressure transition temperature. In most
combined cycle designs, the temperature of the GT exhaust is below this transition temperature.
Due to the importance of this regime, we shall illustrate how to algebraically calculate the HP and IP steam
generation rates for the arrangement of Fig. 1. For clarity, blowdown and minor losses are ignored in the
algebra, all though they are included in the numerical results shown.
The gas available enthalpy drop, from inlet to the HP pinch-point governs the HP steam production rate, i.e.
 (hg ,1 − hg , HPpp ) 
ms , HP = m g 

 (hs , HP − hee, HPE ) 
............................................. (1)
The gas available enthalpy drop from the HP pinch-point to the IP pinch-point must heat the HP water from
the exit of the combined economiser to the exit of the HP economiser. The heat left over after this duty
generates the IP steam. Thus, the IP steam production rate is given by the energy balance:
ms , IP =
m g (hg , HPpp − hg , IPpp ) − ms , HP (hee, HPE − hee,CE )
(hs , IP − hee,CE )
......... (2)
Once the HP steam production rate is computed by Equation (1), the IP steam production rate can be
computed by Equation (2). An increase in the inlet gas temperature will result in more HP steam generation,
but less IP steam generation. This is because the increase in water to be heated in the HP economiser ahead
of the IP evaporator reduces the numerator of Equation (2). However, the increase in HP heat recovery will
invariably be greater than the decrease in IP heat recovery, because the ratio
(hs , HP − hee, HPE )
(hee, HPE − hee,CE )
is invariably greater than one, causing the drop in gas temperature available to the IP to be less than the
increase in inlet gas temperature available to the HP.
With both HP and IP steam generation rates found, the gas enthalpy at the stack may then be found, by an
energy balance over the combined economiser section, in which all quantities are known except hg , stk , i.e.
(ms , HP + ms , IP )(hee,CE − h fw ) = m g (hg , IPpp − hg , stk )
.................. (3)
Finally, the stack temperature Tg , stk is found from the stack specific enthalpy hg , stk .
Figure 3 shows the effect on the T-Q% diagram of raising the initial gas temperature in the pinchconstrained regime. The increase in total HP+IP steam generation increases the flow of water through the
combined economiser below the IP pinch (at the lower left of the diagram), lowering the stack temperature.
In effect, the line representing the gas temperature profile pivots about the IP pinch on the T-Q% diagram.
Since the IP pinch is nearer the stack, the effect of Tg,1 on stack temperature is weaker than for single10-3
 M. A. Elmasri, 1990-2002
Chapter 10: Dual-Pressure HRB/CC Thermodynamics
pressure cases, where the pivot point is higher. For the example of Fig. 3, the proportion of the heat below
the pivot point is only 18%. Thus, increasing Tg,1 by 10° lowers Tstack by only 10° x 18/82 ≈ 2°. By
contrast, in typical single-pressure boilers, raising Tg,1 by 10° lowers Tstack by about 5°.
Dual-Pressure Pinch-Constrained HRB Design - Effect of Tg1 on T vs. %Q Profile
(steam @ 80 bar/500 C & 10 bar/205 C - 1160 psia/932 F & 145 psia/401 F)
1292 F
700
Raising Tg1 depresses Tstk
Temperature, C
600
1112
500
932
400
752
300
572
200
392
100
212 F
0
0
10
20
30
40
50
60
70
80
90
100
Heat Transfer, Q %
Figure 3. T-Q% temperature profile for a Dual-Pressure boiler in the Pinch-Constrained regime
The
minimum
stack
temperature is attained when
the inlet gas temperature is at
the dual-pressure transition
temperature. This depends on
feedwater temperature and
steam conditions.
For our
numerical example, it is 630 °C
(1166 °F), and the temperature
profile at this inlet gas
temperature is shown in Fig. 4.
Figure 4.
Example of a
temperature profile for a DualPressure boiler with the inlet
gas
at
the
Dual-Pressure
Transition
Temperature,
at
which all three constraints,
minimum
stack
and
two
minimum pinches are satisfied.
Min ∆Tpinches of 15 C
both achieved
Min Tstk of 90 C
achieved
10-4
 M. A. Elmasri, 1990-2002
Chapter 10: Dual-Pressure HRB/CC Thermodynamics
With the inlet gas temperature at or near the dual-pressure transition value, the IP steam generation is
typically much smaller than the HP steam generation, raising the issue as to whether the extra complexity is
justifiable. In the present numerical example, at the condition shown in Fig. 4, HP steam flow is 17% of the
flue gas and IP flow only 1.4% of the flue gas, i.e. the IP steam flow is only 8% of the HP flow, besides
being of less value in producing power due to its lower pressure.
1.d. The Stack-Constrained Dual-Pressure Regime:
If the gas were above the dual-pressure transition temperature, the assumption that both pinches are at their
minimum results in a stack temperature below the allowable. Setting the stack temperature at its minimum
allowable value then sets the total heat recovery. Since a given, total heat recovery could be distributed in
any proportion between HP and IP, another criterion needs to be introduced to get an unique solution.
Assuming that the object of the boiler is to power a steam turbine, then it is natural to try to maximise HP,
rather than IP steam generation. Thus, in this regime, we assume that the HP pinch is at its minimum value,
maximising HP steam production, which is found from the energy balance between gas inlet and HP pinch
(just as for a single-pressure system). The IP steam flow can then be found from the overall energy balance
and the known HP steam flow. The corresponding IP pinch, which shall exceed its minimum value, can
then be determined. Within this regime, raising the inlet gas temperature allows greater HP steam flow, but
since total heat recovery is limited by the stack, it results in less IP steam flow.
Figure 5 shows the T-Q profile for our
numerical example in this regime, with
an inlet gas temperature of 660 °C
(1220 °F). The HP pinch is at its
minimum of 15 °C (27 °F) and the
stack is at its minimum of 90 °C (194
°F). The IP pinch, however, has had to
be increased above its minimum, to
20.5 °C (37 °F) to further reduce its
heat recovery, to prevent the stack from
falling below its minimum. Such a
design is of questionable practicality,
since the IP steam production is only
3% of the HP steam production.
Figure 5. Example of a temperature profile for a Dual-Pressure boiler in the Stack-Constrained DualPressure regime. With the inlet gas above the Dual-Pressure Transition Temperature, two constraints, the
HP pinch and the minimum stack are met, but the IP pinch is above its minimum value.
The maximum inlet gas temperature for this regime is the value at which the HP pinch-point temperature
difference and the stack temperature are both at their minimum values and the IP flow is zero. We call this
the single-pressure transition temperature. It is the same value as for a single-pressure boiler where the HP
10-5
 M. A. Elmasri, 1990-2002
Chapter 10: Dual-Pressure HRB/CC Thermodynamics
is the only pressure (as shown in Fig. 6 of the single-pressure chapter). Figure 6 shows the temperature
profile for our numerical example at an inlet gas temperature of 680 °C, just below this transition. The IP is
superfluous at this design condition, since the heat recovery down to the minimum stack temperature can all
be accomplished to the HP.
Figure 6. Example of a temperature profile for a Dual-Pressure boiler in the mixed regime, with the inlet gas
approaching the Single-Pressure Transition Temperature. The IP has effectively vanished and is superfluous.
1.e. The Stack-Constrained Single-Pressure Regime:
If the inlet gas were above the single-pressure transition temperature, then the IP steam production would be
zero, and the HP pinch would be larger than its minimum. The total heat recovery from cooling the gas all
the way down to the minimum stack temperature all goes to the HP without attaining its minimum pinch.
The thermodynamic design would simply be just like a stack-constrained single-pressure boiler, described in
the previous chapter.
Figure 7 summarises the various regimes.
10-6
 M. A. Elmasri, 1990-2002
Chapter 10: Dual-Pressure HRB/CC Thermodynamics
Dual-Pressure Heat Recovery Regimes
(example with steam @ 80 bar/500 C & 10 bar/205 C - 1160 psia/932 F & 145 psia/401 F)
Stack Temperature & Pinch Temp. Difference (C)
120
100
Stack
HP Pinch
IP Pinch
80
60
Pinch- Constrained
StackConstrained
Dual-Pressure
StackConstrained
Single-Pressure
40
20
0
500
550
600
650
700
750
Inlet Gas Temperature (C)
Figure 7. Example illustrating the three constraining regimes in Dual-Pressure heat recovery
2. THERMODYNAMIC DESIGN TRENDS
2.a. Effect of Steam Pressures on Heat Recovery:
100 bar
50 bar
Figure 8. Example illustrating the effect of the HP on the T-Q diagrams in Pinch-Constrained 2-P heat recovery
10-7
 M. A. Elmasri, 1990-2002
Chapter 10: Dual-Pressure HRB/CC Thermodynamics
In the pinch-constrained regime, the HP pressure has very little influence on the stack temperature. Raising
the HP reduces the HP heat recovery and steam generation, leaving more heat to reach the IP and increases
IP steam generation. The effect on total heat recovery is therefore quite weak, as shown by Figure 8, which
illustrates T-Q diagrams of two designs that are identical in every respect except for the HP selection.
Figure 9 shows these same diagrams on a T-Q% plot. The stack temperature and hence the total heat
recovered is virtually the same at both HP values, in strong contrast to single-pressure designs.
Dual-Pressure Pinch-Constrained HRB Design - Effect of HP on T vs. %Q Profile
HP steam @ 50 bar (725 psia) and 100 bar (1450 psia), IP steam @ 10 bar (145 psia)
1292 F
700
Raising HP reduces HP flow & increases
IP flow, has negligible effect on Tstk
Temperature, C
600
1112
500
932
400
752
300
572
200
392
100
212 F
0
0
10
20
30
40
50
60
70
80
90
100
Heat Transfer, Q %
Figure 9. Example illustrating the effect of the HP on the T-Q% diagram in Pinch-Constrained 2-P heat recovery
20 bar
5 bar
Figure 10. Example illustrating the effect of the IP on the T-Q diagrams in Pinch-Constrained 2-P heat recovery
10-8
 M. A. Elmasri, 1990-2002
Chapter 10: Dual-Pressure HRB/CC Thermodynamics
Figure 10 illustrates the effect of raising the IP on the T-Q diagram. As would be expected, raising the IP
reduces its steam generation and raises the stack. Figure 11 shows the same results on a T-Q% diagram. It
should be noted that changing the IP has no effect on HP steam production. The HP evaporator line
lengthens on Fig. 11 because HP heat recovery increases as a percentage of the total, not in absolute amount.
Dual-Pressure Pinch-Constrained HRB Design - Effect of IP on T vs. %Q Profile
HP steam @ 100 bar (1450 psia), IP steam @ 5 and 20 bar (73 and 290 psia)
1292 F
700
Raising IP has no effect on HP flow but reduces
IP flow & raises Tstk (HP flow in % increases)
Temperature, C
600
1112
500
932
400
752
300
572
200
392
100
212 F
0
0
10
20
30
40
50
60
70
80
90
100
Heat Transfer, Q %
Figure 11. Example illustrating the effect of the IP on the T-Q% diagram in Pinch-Constrained 2-P heat recovery
2.b. Effect of Inlet Gas Temperature on Heat Recovery:
Figure 12 shows the variations in stack temperature and pinch points as a function of inlet gas temperature
for several combinations of HP and IP. The figures also show the single-pressure results, where the single
pressure is the same as the HP of the dual pressure cases.
The figures show that the dual-pressure transition temperature depends primarily on the IP and only weakly
on the HP. For a 5-bar (73 psia) IP, it can be in the 620-640 °C (1150-1185 °F) range, just beyond the upper
bound of the exhaust temperature of the most modern gas turbines.
Figure 12a shows the value of the second pressure is substantial when the flue gases are at a modest
temperature. With the cooler exhaust temperatures of aeroderivatives, a dual-pressure boiler can lower stack
temperature by about 75°C (135 °F) relative to a single pressure boiler. With the hotter gases of a heavy
duty GT, the dual-pressure boiler can lower the stack by about 50 °C (90 °F) relative to a single-pressure
design. The second pressure is advantageous only when single-pressure heat recovery is pinch-constrained,
not stack constrained. If the initial gas temperature, Tg1, is high enough for the minimum stack to be
attained, even with a single pressure boiler at the desired HP, a multi-pressure system would have no benefit
since all the heat that can be recovered can be at HP.
10-9
 M. A. Elmasri, 1990-2002
Chapter 10: Dual-Pressure HRB/CC Thermodynamics
Dual-pressure & Single-pressure Heat Recovery: Stack Temperature
842 F
932
1022
1112
1202
240
1472
1562 F
464 F
1-P 100 bar (1450 psia)
220
Stack Temperature (C)
1382
1292
428
100/20 bar (1450/290 psia)
100/10 bar (1450/145 psia)
200
392
1P 50 bar (725 psia)
50/10 bar (825/145 psia)
180
356
50/5 bar (825/83 psia)
160
320 F
140
284
120
248
100
212
80
450
500
550
600
650
700
750
800
850
Main Assumptions:
Min. pinch = 15 C (27 F)
Min. stack = 90 C (194 F)
Appr. Subcooling=5 C (9 F)
Tfw=33 C (91 F)
Tstm=540 C but <=Tg-20 C
(1004 F but <=Tg-36 F)
176 F
Exhaust Gas Temperature (C)
Dual-pressure & Single-pressure Heat Recovery: Pinches
842 F
932
Pinch Temperature Difference (C)
80
1022
1112
1202
1382
1292
1472
1562 F
144 F
HP Pinch, 100 bar (1450 psia)
70
126
IP Pinch, 100/20 bar
IP Pinch, 100/10 bar
60
108
HP Pinch, 50 bar (725 psia)
IP Pinch, 50/10 bar
50
90
IP Pinch, 50/5 bar
40
72 F
30
54
20
36
10
450
500
550
600
650
700
750
800
850
Main Assumptions:
Min. pinch = 15 C (27 F)
Min. stack = 90 C (194 F)
Appr. Subcooling=5 C (9 F)
Tfw=33 C (91 F)
Tstm=540 C but <=Tg-20 C
(1004 F but <=Tg-36 F)
18 F
Exhaust Gas Temperature (C)
Figures 12a & 12b. Stack Temperature (12a-top) and Pinch Point Temperature Differences (12b-bottom) as functions of flue
gas inlet temperature at different steam pressures
10-10
 M. A. Elmasri, 1990-2002
Chapter 10: Dual-Pressure HRB/CC Thermodynamics
Dual-pressure & Single-pressure Heat Recovery: Steam Generation
932
1022
1112
1202
1382
1292
1472
1562 F
5
25
4
20
HP Flow, 100 bar (1450 psia)
3
15
IP Flow, 100/20 bar
IP Flow, 100/10 bar
HP Flow, 50 bar (725 psia)
2
10
IP Flow, 50/10 bar
IP Flow, 50/5 bar
1
5
0
450
500
550
600
650
700
750
800
HP Steam Production, % of Gas Flow
IP Steam Production, % of Gas Flow
842 F
Main Assumptions:
Min. pinch = 15 C (27 F)
Min. stack = 90 C (194 F)
Appr. Subcooling=5 C (9 F)
Tfw=33 C (91 F)
Tstm=540 C but <=Tg-20 C
(1004 F but <=Tg-36 F)
850
Exhaust Gas Temperature (C)
Figure 13. HP and IP steam generation rates as functions of flue gas inlet temperature at different HP/IP steam pressures
Dual-pressure & Single-pressure Combined Cycle: Heat Recovery Efficiency
842 F
932
1022
1112
1202
95
1292
1382 F
90
HRB Efficiency, %
85
80
Main Assumptions:
Min. pinch = 15 C (27 F)
Min. stack = 90 C (194 F)
Appr. Subcooling=5 C (9 F)
Tfw=33 C (91 F)
Tstm=540 C but <=Tg-20 C
(1004 F but <=Tg-36 F)
75
70
100 bar (1450 psia)
100/20 bar
100/10 bar
50 bar (725 psia)
50/10 bar
50/5 bar
65
60
55
50
45
450
500
550
600
650
700
750
Exhaust Gas Temperature (C)
Figure 14. Heat Recovery Efficiency as a function of flue gas inlet temperature at different HP/IP steam pressures
10-11
 M. A. Elmasri, 1990-2002
Chapter 10: Dual-Pressure HRB/CC Thermodynamics
Figure 13 illustrates the principle that any design parameter that tends to increase HP heat recovery will tend
to decrease IP heat recovery and vise versa. The maximum IP flows are at the HP conditions which
minimise HP heat recovery, viz high HP and low inlet gas temperature, and vise versa.
For instance, if one tries to make 100 bar (1450 psia) steam from an unfired aeroderivative exhaust at 450
°C (842 °F), the HP steam generation would be about 7.5% of the flue gas. Thus, the IP steam generation, at
10 bar (145 psia) say, would be quite high, at about 4% of the flue gas. The IP steam flow is thus more than
half the HP steam flow. By contrast, if one generates HP steam at a modest 50 bar (725 psia) from the hot
exhaust of a heavy-duty GT at 550 °C (1022 °F), the HP steam generation would be about 13% of the gas
flow, and the IP steam generation would be just 2% of the gas flow, i.e. about 1/6th of the HP.
Figure 14 shows how heat recovery efficiency is greatly improved by the second pressure when the inlet gas
temperature is low, but that the second pressure becomes less useful at higher inlet gas temperatures.
2.c. Dual-Pressure Bottoming Cycle Performance:
Dual-pressure Combined Cycle: ST Output Relative to Single-pressure
842 F
ST Output, Relative to Single-pressure
1.5
932
1022
1112
1202
1292 F
100/20 bar
100/10 bar
1.4
Main Assumptions:
Single-pressure=HP of Dualpressure
Min. pinch = 15 C (27 F)
Min. stack = 90 C (194 F)
Appr. Subcooling=5 C (9 F)
Tfw=33 C (91 F)
Tstm=540 C but <=Tg-20 C
(1004 F but <=Tg-36 F)
Pcond=50 mb (0.725 psia)
50/10 bar
50/5 bar
1.3
1.2
1.1
1
450
500
550
600
650
700
Exhaust Gas Temperature (C)
Figure 15. Illustration of the Advantage of a Dual-pressure bottoming cycle over a Single-pressure design
Figure 15 shows how the second pressure increases bottoming cycle output relative to a single-pressure
cycle. With the cooler exhaust of aeroderivatives, that improvement is on the order of 20%, whereas with
the hotter exhaust of a typical heavy duty GT, the improvement is on the order of 10%. Please note that in
Fig. 15, the improvement is shown relative to a single-pressure cycle designed at the HP of the alternative
dual-pressure cycle, so it is overstated for the 100 bar designs, especially at modest inlet gas temperatures,
since a single-pressure design should not be considered at 100 bar, except for a larger plant with a high inlet
gas temperature, typically associated with supplementary-fired designs.
The increased ST output is less than proportionate to the increased heat recovery, however, since the
10-12
 M. A. Elmasri, 1990-2002
Chapter 10: Dual-Pressure HRB/CC Thermodynamics
additional heat recovered is of low-grade, with a relatively low exergy and generates IP steam, whose work
per unit of flow is less than for HP steam. This can be seen by comparing Figures 16 and 15. Thus, the
dual-pressure bottoming cycle exhibits lower internal efficiency than a single-pressure cycle, although its
overall efficiency is higher. Figs. 17 and 18 show these effects.
Dual-pressure HRB: Heat Recovery Relative to Single-pressure
Heat Recovered, Relative to Single-pressure
842 F
932
1022
1.5
1112
1202
1292 F
100/20 bar
100/10 bar
1.4
50/10 bar
Main Assumptions:
Single-pressure=HP of Dualpressure
Min. pinch = 15 C (27 F)
Min. stack = 90 C (194 F)
Appr. Subcooling=5 C (9 F)
Tfw=33 C (91 F)
Tstm=540 C but <=Tg-20 C
(1004 F but <=Tg-36 F)
50/5 bar
1.3
1.2
1.1
1
450
500
550
600
650
700
Exhaust Gas Temperature (C)
Figure 16. Illustration of the increased heat recovery of a Dual-pressure HRB over a Single-pressure design
Dual-pressure & Single-pressure Combined Cycle: Steam Cycle Internal Efficiency
842 F
Steam Cycle Internal Efficiency, %
37
932
1022
1112
1202
1292
1382 F
36
35
Main Assumptions:
Min. pinch = 15 C (27 F)
Min. stack = 90 C (194 F)
Appr. Subcooling=5 C (9 F)
Tfw=33 C (91 F)
Tstm=540 C but <=Tg-20 C
(1004 F but <=Tg-36 F)
Pcond=50 mb (0.725 psia)
34
33
100 bar (1450 psia)
100/20 bar
100/10 bar
50 bar (725 psia)
50/10 bar
50/5 bar
32
31
30
29
450
500
550
600
650
700
750
Exhaust Gas Temperature (C)
Figure 17. Comparison of Steam Cycle Internal Efficiencies for Dual-pressure and Single-pressure designs
10-13
 M. A. Elmasri, 1990-2002
Chapter 10: Dual-Pressure HRB/CC Thermodynamics
Dual-pressure & Single-pressure Combined Cycle: Steam Cycle Overall Efficiency
842 F
932
1022
1112
1202
1292
1382 F
Steam Cycle Overall Efficiency, %
32
30
Main Assumptions:
Min. pinch = 15 C (27 F)
Min. stack = 90 C (194 F)
Appr. Subcooling=5 C (9 F)
Tfw=33 C (91 F)
Tstm=540 C but <=Tg-20 C
(1004 F but <=Tg-36 F)
Pcond=50 mb (0.725 psia)
28
26
100 bar (1450 psia)
24
100/20 bar
100/10 bar
22
50 bar (725 psia)
50/10 bar
20
18
50/5 bar
450
500
550
600
650
700
750
Exhaust Gas Temperature (C)
Figure 18. Comparison of Bottoming Cycle Overall Efficiencies for Dual-pressure and Single-pressure designs
2.d. Selecting Dual-Pressure Bottoming Cycle Pressures:
The performance of dual-pressure bottoming cycles is less sensitive to the selection of pressures than for
single-pressure systems. As a rule, smaller turbines need lower pressures, to avoid excessively short blading
and hence lower blading efficiency. Figures 19 & 20 show the influence of the HP and IP pressure
selections on the thermodynamic performance of a small (15 MW class) and a medium (100 MW class)
bottoming cycle. The figures also show, for comparison purposes, the corresponding single-pressure curve
from Fig. 21 of the previous chapter.
The thermodynamically optimum HP is in the range 40-60 bar (580-870 psia) for a small (15 MW) dualpressure bottoming cycle, and 80-100 bar (1150-1450 psia) for a medium (100 MW) cycle. The higher end
of these ranges give the best thermodynamics, but may produce excessive moisture in the last turbine stages
if the condenser pressure is low, as illustrated in Fig. 20 of the previous chapter. Higher HP, in the range
mentioned, increases plant cost more than the modest increase in output, thereby increasing the specific cost
($/kW). Thus, the practical optimum value of the HP is generally lower than the thermodynamic optimum,
being in the 40 bar (600 psia) range for a small (15 MW class ST) cycle and in the 70 bar (1100 psia) range
for a medium (100 MW class ST).
The thermodynamically optimum IP is in the range 3-5 bar (45-75 psia) for a small (15 MW) dual-pressure
bottoming cycle, and 4-8 bar (60-120 psia) for a medium (100 MW) cycle. The lower values result in low
stack temperatures, which necessitate taller stacks, increase possibility of corrosion, and demand stronger
measures to raise stack temperature when oil is used as a backup fuel. In addition, the increased heat
recovery requires a larger boiler. The larger mass flow of low-pressure, high-volume steam increases IP
piping cost. The larger flow through the ST exhaust requires longer blades. The increased heat rejection at
10-14
 M. A. Elmasri, 1990-2002
Chapter 10: Dual-Pressure HRB/CC Thermodynamics
the condenser requires a larger condenser and CW pumps. Thus, the economically optimum IP is higher
than the thermodynamic optimum, since the cost savings with a higher IP will outweigh the minor sacrifice
in performance. This leads to ‘optimum’ IP selections in the 5-10 bar (75-150 psia) range.
Bottoming Steam Cycle Overall Efficiency, %
Bottoming Cycle Pressure Optimisation, 15 MW Class ST
28.0
27.5
27.0
3 bar (44 psia) IP
26.5
6 bar (87 psia) IP
26.0
9 bar (131 psia) IP
25.5
Single-pressure
25.0
Exhaust Gas:
100 kg/s (220 lb/s)
550 C (1022 F)
24.5
24.0
20
40
60
80
Throttle Pressure (bar)
Figure 19. Effect of Steam Pressures on Overall Efficiency of Small Bottoming Cycles
Bottoming Steam Cycle Overall Efficiency, %
Bottoming Cycle Pressure Optimisation, 100 MW Class ST
28.5
28.0
4 bar (58 psia) IP
27.5
8 bar (116 psia) IP
27.0
12 bar (174 psia) IP
26.5
Single-pressure
26.0
25.5
Exhaust Gas:
600 kg/s (1320 lb/s)
550 C (1022 F)
25.0
24.5
20
40
60
80
100
Throttle Pressure (bar)
Figure 20. Effect of Steam Pressures on Overall Efficiency of Medium Bottoming Cycles
Another consideration in selecting the IP, is that it can be a convenient source of steam injection for NOx
control. The amount typically needed is approximately equal to IP steam production. NOx steam needs to
be at a pressure 30-35% higher than GT compressor discharge, which means 15-25 bar (220-360 psia) for a
10-15
 M. A. Elmasri, 1990-2002
Chapter 10: Dual-Pressure HRB/CC Thermodynamics
heavy duty gas turbine. This is higher than thermodynamic optimum for good heat recovery but makes
practical sense in many cases. For a high pressure ratio aeroderivative, any NOx steam needed may be at as
much as 650 psia (45 bar), which is much too high for an IP. Such steam would typically be bled from the
HPT or let-down from the HP steam header.
2.e. Supplementary Firing in Dual-Pressure Designs:
Due to the more effective heat recovery with dual-pressure designs, pinch-constrained heat recovery
efficiency is less sensitive to inlet gas temperature than a single-pressure system. As illustrated by Figure 3,
the effect of increasing inlet gas temperature by δTg,1 on lowering stack temperature by δTstack is much
weaker than for single-pressure designs.
Typically, (δTstack/δTg,1)Dual-pressure ≈ 0.15-0.25, whereas (δTstack/δTg,1)Single-pressure ≈ 0.3-0.5.
Thus, supplementary firing will benefit the dual pressure HRB about half as much as the single pressure
HRB, and the Duct Burner Fuel Utilisation efficiency will be in the 115%-125% range.
Dual pressure combined cycles will be less-efficient with supplementary firing since, in addition to the
lower DBFU efficiency, the steam cycle internal efficiency is lower than a single-pressure cycle.. Thus, the
efficiency of converting duct burner fuel to power is about
η DB _ fuel → Power = η DBFU ∗ηi ≈ 1.2 * 0.3 ≈ 0.36 (36%)
which is much less than the base efficiency of a typical dual-pressure CC.
2.f. HRB Surface Area for Dual-Pressure Designs:
Boiler Area, Relative to Single-pressure
Dual-pressure Required Boiler Area, Relative to Single-pressure
842 F
2
932
1022
1112
1202
1292 F
100/20 bar
1.9
100/10 bar
1.8
50/10 bar
Main Assumptions:
Single-pressure=HP of Dualpressure
Min. pinch = 15 C (27 F)
Min. stack = 90 C (194 F)
Appr. Subcooling=5 C (9 F)
Tfw=33 C (91 F)
Tstm=540 C but <=Tg-20 C
(1004 F but <=Tg-36 F)
DPgas=20 mb (8" H2O)
50/5 bar
1.7
1.6
1.5
1.4
1.3
1.2
1.1
1
450
500
550
600
650
700
Exhaust Gas Temperature (C)
Figure 21. Typical Required Boiler Surface Area for a Dual-Pressure HRSG relative to a Single-Pressure Design
10-16
 M. A. Elmasri, 1990-2002
Chapter 10: Dual-Pressure HRB/CC Thermodynamics
Figure 21 shows that specific HRB surface areas for dual-pressure are much larger than for single-pressure,
shown on Fig. 26 of the single-pressure chapter. With low exhaust gas temperatures, typical of
aeroderivatives, the dual-pressure HRB requires 60-80% more area (but recovers only 25-40% more heat,
per Fig. 16). With the exhaust temperatures typical of heavy-duty gas turbines, the dual-pressure HRB
requires 40-60% more area (but recovers only 10-20% more heat, per Fig. 16).
3. EFFECT OF KEY PARAMETERS
3.a. Sensitivity to Pinches (in the Pinch-Constrained Regime):
Increasing the HP pinch reduces HP steam production, just as for a single-pressure boiler. This lets hotter
gas reach the IP, increasing its steam flow. The increase in IP flow is typically greater than the decrease in
HP flow. This is because the HP flow is set by the enthalpy difference between final HP steam and
feedwater to the HP evaporator, which is usually greater than the enthalpy difference between IP steam and
feedwater to its evaporator. The effect on stack temperature is negligible, falling very slightly if the increase
in IP steam exceeds the decrease in HP steam. Dual-pressure CC output falls slightly, since IP steam is
substituted for HP steam.
As a rough rule, increasing the HP pinch by 5 °C (9 °F) will reduce HP steam output by about 2%. The IP
steam output will increase by a roughly equivalent (but slightly larger) amount in kg/s (lb/s). The ST output
will fall by about 0.2%.
IP pinch does not influence HP flow. Increasing it reduces IP flow, CC power output, and raises stack
temperature. The effect of IP pinch on power is stronger than that of HP pinch, since IP is the last chance to
capture heat before the stack, whereas heat unrecovered at HP is captured at IP.
As a rough rule, increasing the IP pinch by 5 °C (9 °F) will reduce IP steam output on the order of 10% (the
effective gas ∆T for the IP is much smaller than for the HP, and is on the order of 1/5th of that for the HP).
The ST output will fall by about 0.3%.
Decreasing either pinch increases power output linearly, but increases the required HRB area exponentially.
It is thus uneconomical to reduce a pinch below some reasonable value, like 10 °C (10 °F).
3.b. Sensitivity to Approach Subcooling:
Increasing approach subcooling reduces steam production, much like increasing a pinch. Thus, the effect of
changing approach subcoolings by a certain number of degrees is very similar to, but weaker than, the effect
of changing pinches by the same number of degrees.
As a rough rule, increasing the HP approach subcooling by 5 °C (9 °F) will reduce HP steam output by
about 1%. The IP steam output will increase by a roughly equivalent (but slightly larger) amount in kg/s
(lb/s). The ST output will fall by about 0.1%. No similarly general rule can be made for the IP subcooling,
since its impact on IP steam production depends, amongst other things, on whether the HP economiser
upstream (in the gas path) of the IP evaporator receives water directly from the IP economiser exit.
3.c. Sensitivity to Steam Temperatures:
Reducing HP steam temperature would increase HP steam production, since less heat would be absorbed in
the HP superheater, leaving more for the HP evaporator. The increased HP water flow would absorb more
10-17
 M. A. Elmasri, 1990-2002
Chapter 10: Dual-Pressure HRB/CC Thermodynamics
heat in the HP economiser, leaving a lower temperature gas for the IP section, whose flow would decline.
The reduced HP steam temperature reduces the percentage of HP enthalpy gained above the pinch to that
below the pinch. The shift in balance towards a relatively greater requirement for low grade heat results in a
cooler stack, so the total heat recovery increases.
The increased HP steam flow and lower stack temperature do not, however, mean more power. Since the
increased HP flow is at a lower temperature and enthalpy, the expansion line to the condenser is shorter, and
the drop in work per unit of steam mass flow outweighs the higher mass flow. This, together with the
reduced IP flow result in a loss of power output. The cycle energy balance would show a reduction in stack
loss but a greater increase in condenser loss.
As a rough rule, reducing HP steam temperature by 10°C (18 °F) will reduce the stack temperature by ½ °C
(1 °F), will increase HP steam output by 1%, and will decrease ST power output by 0.15%.
Reducing IP steam temperature has the same general effects as for HP temperature, but less impact on
overall performance, since the IP role is secondary.
3.d. Sensitivity to Condenser Pressure:
Lowering the condenser pressure has a strong effect on output. It also necessitates more condenser surface
for a given cooling water initial temperature and rise, since the mean condenser temperature difference is
reduced. The lower condenser temperature also reduces initial feedwater temperature, necessitating larger
economiser surface areas in the HRB and reducing stack temperature.
The additional condenser and HRB surface areas needed are insignificant compared to the power and
efficiency gain, however, the increase in size of the last steam turbine stages is a significant cost factor.
Another limitation to reduction of condenser pressure comes from the reduced ST exhaust quality, which
drops (wetter steam) with condenser pressure.
As a rough rule, a 20 mb (0.29 psia or 0.58” Hg) reduction in condenser pressure increases ST output by
about 1%.
3.e. Sensitivity to Steam Turbine Efficiency:
An increase in overall isentropic efficiency of the ST by 1% will also increase power output by 1%.
However, an increase of 1% in basic stage efficiency will have a much smaller impact for two reasons:
(a) The so-called ‘reheat effect’: Increasing the efficiency of each stage delivers cooler steam to the next.
Thus the overall isentropic efficiency obtained by compounding many stages of a certain efficiency is less
than the efficiency of the stages.
(b) The moisture effect: A more efficient turbine results in a lower exhaust quality, with greater moisture in
the low pressure stages. Moisture reduces the efficiency of these stages to below its basic value, realised
with dry steam. Thus, depending on the proportion of the expansion line in the wet zone, the impact of a
more efficient turbine will be diluted.
As a rough rule, in a typical dual-pressure cycle, a 1 percentage point improvement (from 87% to 88%, say)
in basic (dry) stage efficiency will result in a 2/3rds of 1% gain in ST output.
10-18
 M. A. Elmasri, 1990-2002
Chapter 10: Dual-Pressure HRB/CC Thermodynamics
3.f. Sensitivity to HRB Draft Loss:
Designing with a higher boiler gas-side pressure loss allows the boiler to be made smaller, with higher gas
velocities and heat transfer coefficients (within the reasonable range).
The greater draft loss curtails GT power output, but causes the GT exhaust temperature to climb. The latter
effect increases steam production and ST power. Depending on the GT correction curves, the change in ST
power may substantially compensate for the change in GT power, being comparable in magnitude and
opposite in effect, resulting in a negligible impact on plant net power. In such cases, however, plant net heat
rate will still suffer with a higher draft loss. For many gas turbines, however, the change in ST power would
compensate for just half the change in GT power, causing the total CC power to change by roughly half the
change in GT power.
Higher draft losses allow reducing HRB surface area. As a rough rule, a 5 mb (2” H2O) increase in draft
loss enables a 3-6% reduction in boiler surface area, and a 1½-3% reduction in boiler cost.
4. CONVENTIONAL STEAM PLANT OR BOTTOMING CYCLE ?
In our previous discussions, we saw that heat recovery from a hot gas to generate steam can be pinchconstrained or stack-constrained. At a certain hot gas temperature, which we termed the transition
temperature, the two constraints were applicable simultaneously.
For a single-pressure boiler, we can think of the heat needed by the water/steam as being of two categories:
(i) high-grade heat, which is needed to evaporate and superheat the steam (at temperatures above the pinch),
and (ii) low-grade heat, which is needed to heat the water to (near) saturation (at temperatures below the
pinch). If the inlet gas is at the transition temperature, the proportion of its total available heat which is of
high grade (above the pinch) and that which is of low grade (below the pinch), are perfectly matched to the
needs of the water/steam for each category of heat.
If the gas inlet temperature is lower than transition, the gas available heat, from inlet to minimum stack, has
too small a proportion of high-grade heat, so its full energy cannot be used. The pinch constrains the heat
recovered to be commensurate with the available high grade heat, and the surplus low grade heat is wasted
(the stack is warmer than its minimum value). The remedy for this problem in a combined cycle was to
introduce additional evaporators at lower pressures to harness the surplus low grade heat, add the resulting
steam to the ST at lower pressures, and make additional power from it.
If the gas inlet temperature is higher than transition, the opposite happens. The gas available heat, from inlet
to minimum stack, has too high a proportion of high-grade heat, so its full exergy cannot be used. The stack
constrains the total heat recovered, and a portion of the high grade heat, which is at a temperature high
enough to evaporate and superheat steam, is used instead to heat water, wasting exergy.
Therefore, if the heat source temperatures is in excess of the single-pressure transition value, such as from
conventional furnaces, regenerative feedwater heating is used to improve steam cycle efficiency. A
conventional steam plant with steam extraction from the turbine to heat feedwater is the logical opposite of a
multi-pressure bottoming cycle.
The conventional steam plant furnace provides so much high-grade energy, that it would be a waste to use it
to heat low-temperature water, as illustrated in Figure 22. The heat above saturation temperature by some
minimum, practical pinch difference, is best utilised making high-pressure, high-temperature steam (or
10-19
 M. A. Elmasri, 1990-2002
Chapter 10: Dual-Pressure HRB/CC Thermodynamics
reheating it). To save more of it for those high-temperature purposes, feedwater is heated by lower-grade
steam extracted from the turbine after partial expansion from the throttle. Flue gas low grade heat is then
recycled to save fuel (high-grade heat) in an air preheater, where each BTU or kJ transferred to the
combustion air saves one BTU or kJ of fuel, i.e. one BTU or kJ of high-grade heat.
The typical bottoming cycle, on the other hand, suffers from a scarcity of high-grade but an excess of lowgrade heat. Thus it benefits from "negative regenerative feed heating" i.e. multi-pressure steam generation
with the low-grade steam added to the steam turbine to complement the expansion to the condenser.
Thus with initial gas temperatures low enough for heat recovery to be pinch-controlled (typical unfired GT
exhaust), regenerative heating has a negative impact and its opposite, multi-pressure heat recovery, is used
to increase efficiency. With intermediate gas temperatures, in the range where the pinch and stack
constraints are roughly equivalent (characterised by the economiser and gas temperature profiles being
roughly parallel), neither multi-pressure recovery nor regenerative feed heating should be applied. At high
gas temperatures, the heat recovery is stack-controlled, and regenerative heating should be used rather than
multi-pressure heat recovery.
T (C)
1400
1200
1000
800
600
High grade heat,
can generate or
superheat steam,
not merely heat
water
400
200
Q (%)
Figure 22. Temperature profile in a high temperature boiler, showing that the proportion of high grade
heat available from the heat source exceeds that needed by the water/steam temperature profile.
Regenerative feedwater heating should be employed under such conditions.
10-20
 M. A. Elmasri, 1990-2002
Chapter 11: Off-Design Behaviour of HRSG’s
OFF-DESIGN BEHAVIOUR OF HEAT RECOVERY BOILERS
Revised September, 2002
© Maher Elmasri 1990-2002
As in previous chapters, the single-pressure heat recovery boiler is used to clearly illustrate the principles.
Once these have been elucidated, it becomes easier to understand the more complex behaviour of multipressure boilers.
11.1
OFF-DESIGN PERFORMANCE OF A SINGLE-PRESSURE HRSG
To illustrate the principles, we first consider a simple, single pressure heat recovery boiler consisting of a
main economiser, circulating-drum evaporator and a superheater. For a practical example, we also include
an Integral Deaerator within the feedwater heating section of the HRSG, since this is quite common in
practice. The flow schematic at the nominal design condition is shown in Figure 1a.
59 T
12.75 M
156 T
24.61 p
239 T
V4
900 p 863 T 12.74 M
71.14 %N2
14 %O2
3 %CO2+SO2
11 %H2O
0.8568 %Ar
LTE
LPB
IPE2
IPB
IPS2
24.61 p
156 T
12.75 M
24.61 p
239 T
1.125 M
930.6 p
515 T
12.75 M
930.6 p
536 T
12.74 M
927.8 p
867 T
12.74 M
337 T
100 M
1000 T
100 M
384
425
568
892
1000
p[psia], T[F], M[lb/s], Steam Properties: IAPWS-IF97
0 09-28-2002 07:51:50 file=C:\SEMINAR\Sem_5-02\Sec11_HRSG_OD\1PHRB_9-02.GTM
Figure 1a. Nominal design-point heat & mass balance of the single-pressure HRSG used in the numerical examples
The boiler used in the example is designed for a nominal flue gas mass flow rate of 100 lb/s at a temperature
of 1000 °F (538 °C), the exhaust conditions of a gas turbine in the 12-15 MW class. Nominal steam
conditions at superheater outlet are 900 psia (62 bar) and 863 °F (462 °C). The design is cost-driven, with a
fairly large nominal pinch-point of 32 °F (17.8 °C), and with an economiser approach subcooling of 21 °F
(11.7 °C).
The nominal, design-point temperature profile is shown in Figure 1b. At the cool end of the profile, the
feedwater is first heated in a Low Temperature Economiser (LTE), then delivered to an Integral Deaerator.
The LTE (sometimes referred to as a condensate preheater) heats the feedwater to a few degrees below
saturation temperature at the deaerator pressure. The integral deaerator is essentially an evaporator which
generates steam at low pressure, and uses that steam to heat the incoming, subcooled feedwater to its
saturation temperature, driving out dissolved gases in the process, since the solubility of gases in water
11-1
 M. A. Elmasri, 1990-2002
Chapter 11: Off-Design Behaviour of HRSG’s
approaches zero as the water approaches saturation temperature. All the heat absorbed from the flue gases
to generate steam is redeposited in the feedwater, so that the presence of a horizontal section in the
temperature profile of Fig. 1b affects the temperature differences and the heat transfer rate calculations but
does not affect the energy balance, which is the same as if the water was simply heated in a continuous
economiser, rather than in two distinct economisers with an integral deaerator in between 1.
Net Power 0 kW
LHV Heat Rate 0 BTU/kWh
GT MASTER 10.9 Maher Elmasri
1100
LPB
1071.8
LTE
1229.8
1000
HRSG Temperature Profile
IPE2
3738
IPS2
3019
IPB
8785
6
Q BTU/s
900
10
800
TEMPERATURE [F]
700
600
12
500
14
17
400
300
200
UA BTU/s-F
100
4.884
6.535
0
0
2
65.87
13.32
35.53
4
6
8
10
12
14
16
18
20
HEAT TRANSFER FROM GAS [.001 X BTU/s]
Figure 1b. Nominal design-point temperature profile of the single-pressure HRSG used in the numerical examples
The overall dimensions of such a boiler are shown in Figure 2. It has 38 tube rows, five for the superheater,
seventeen for the evaporator, and eleven for the main economiser. The Integral Deaerator and the Low
Temperature Economiser have two rows and three rows, respectively.
Naturally, the presence of two distinct pumps, one pressurising the water to the deaerator pressure and a second pump
pressurising the water from deaerator pressure to the final boiler pressure has some secondary effects on the heat
balance. Other secondary effects are differences in minor heat losses attributed to the deaerator and its vent condenser.
1
11-2
 M. A. Elmasri, 1990-2002
Chapter 11: Off-Design Behaviour of HRSG’s
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FOR QUALITATIVE INDICATION ONLY
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Thermoflow, Inc.
Company: Thermoflow
User: Maher Elmasri
HEAT RECOVERY STEAM GENERATOR
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Date: 09/30/02
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Drawing No:
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PEACE/GT MASTER 10.9
Figure 2. Schematic showing overall dimensions of the single-pressure HRSG used in the numerical examples
11.1.1 EFFECT OF VARYING BOILER PRESSURE – THE PRESSURE-FLOW CHARACTERISTIC
11.1.1.1 The Pressure-Flow “Pumping Characteristic” of a Heat Recovery Boiler
Viewed as a black-box, a heat recovery boiler along with its controls, appears to the surrounding system as
though it were a pump, with a drooping pressure-mass flow characteristic. Its operating point is determined
by the flow resistance it sees at its outlet, imposed by the surrounding system. Reducing that resistance,
such as by opening a control valve on the receiving steam turbine, lowers pressure and increases mass flow
and vise versa. At each pressure/mass flow combination, the steam temperature leaving the superheater has
a unique value.
The boiler’s overall pumping characteristic should not be confused with that of its feedpump, which is
actually one component within the boiler “black-box”. The feedpump generally has a similar characteristic,
but this is not relevant to the system at large because the feedpump is controlled as a slave to the boiler. If
the feedpump is of the fixed-speed type, its control valve is set to throttle its outlet, wasting some of its
capacity, so that the valve can be modulated to maintain drum water level. If the pump is of the variablespeed type, its speed is regulated to maintain drum water level. Thus, any event which increases steam
generation rate, thereby reducing drum level, causes the pump to deliver more water to restore drum level,
and vise versa. Thus, the pump does not set steam generation rate, but is controlled to supply feedwater in
an amount that replenishes the steam generation rate dictated by the heat transfer surfaces and the hot gas
conditions. For given heat transfer surfaces, it is hot gas conditions: temperature, flow rate, and
composition, which dictate the boiler’s steam generation rate as a function of boiler pressure. This
relationship between steam generation rate and steam pressure constitutes the boiler’s “pumping
characteristic”, seen by the system receiving steam from the boiler, and quite distinct from that of the
boiler’s own feedpump.
11-3
 M. A. Elmasri, 1990-2002
Chapter 11: Off-Design Behaviour of HRSG’s
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Ts - Ts,nominal
Steam Pressure, % of nominal
1-P HRB Pressure-Flow Characteristic
F
Fig. 3 shows the boiler’s pumping characteristic, of the present HRB example. Feedwater is assumed at 59
°F (15 °C), and hot gas flow into the HRB is fixed at 100 units of massflow at 1000 °F (538 °C) with the
composition shown in Fig. 1a. Raising the delivery pressure, such as by closing a control valve, reduces
steam mass flow because the increasing saturation temperature diminishes the available temperature
difference between hot gas and boiling temperature. Diminished steam flow, on its own, would tend to
reduce heat capacity (m Cp) of the steam going through the superheater, causing one to expect the steam
temperature leaving the superheater to rise. However, since the specific heat of steam, Cp, increases with
pressure, the heat capacity (m Cp) of the steam flowing through the superheater rises, rather than drops, in
our example, causing the steam exit temperature to drop as pressure is raised and flow diminished, as seen in
Figure 3. This trend could go in the opposite direction in other numerical examples.
-4
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106
Steam Flow, % of nominal
Nominal Steam Flow = 12.8% of Gas Flow
Nominal Steam Conditions = 900 psia / 864 oF (62 bar / 462 oC)
Gas Temperature = 1000 oF (538 oC)
Steam Pressure
Ts - Ts,nominal
Figure 3. Typical example of the pressure-flow “pumping characteristic” of a single-pressure boiler
11.1.1.2 Effect of Boiler Pressure on Stack Temperature & Integral D/A Pressure
In the example, a floating pressure Integral Deaerator is used, with its evaporator labelled “LPB” in Figures
1a and 1b. In typical practice, this sort of deaerator operates with a floating pressure, set by the balance
between heat transfer rate from the flue gases to the LPB tubes, which sets LPB steam generation rate, and
the rate of condensation of that generated steam by the subcooled, incoming feedwater. If steam generation
rate were to exceed condensation rate, the pressure would rise, raising the LPB saturation temperature, and
bringing the horizontal line representing it on Fig. 1b closer to the line representing flue gas temperature.
This diminished temperature difference between gas and steam would then result in a reduction in heat
transfer rate from the gases, hence in LPB steam generation rate. The increased saturation temperature also
increases the extent of subcooling of the incoming feedwater, which increases steam condensation rate. The
simultaneous reduction in steam generation rate and increase in steam condensation rate restores equilibrium
between the two, but at a higher LPB pressure.
11-4
 M. A. Elmasri, 1990-2002
Chapter 11: Off-Design Behaviour of HRSG’s
In practice, the Integral Deaerator/LPB pressure is left to float, but cannot be allowed to exceed a safe value
for the pressure vessels, so a vent would open to blow off steam if the pressure becomes too high. Similarly,
it cannot be allowed to float to an excessively low pressure, since if it falls below the atmospheric pressure,
the non condensable gases cannot be vented without a vacuum pump. Thus, in many installations, steam
from the main boiler pressure is added to the LPB to "peg it" if its pressure floats too low. In our numerical
example, the pegging set-point is assumed to be 16 psia (1.1 bar).
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1-P HRB Performance
Effect of Steam Pressure on D/A & Stack
C
Figure 4 shows how the stack temperature and D/A pressure change with operating condition. Higher main
steam pressures diminishes evaporation rate in the main evaporator and raises the stack temperature, just as
for a design-point calculation. Since the gas temperature arriving at the LPB increases, the LPB pressure
floats upwards, as described above. Lowering the main pressure has the opposite effect on stack
temperature and D/A pressure. At main steam delivery pressures below 60% of nominal, the D/A pressure
would fall too low and must be pegged in our numerical example.
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D/A Pressure, % of nominal
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Steam Pressure, % of nominal
Nominal Steam Conditions = 900 psia / 864 oF (62 bar / 462 oC)
Nominal D/A Pressure = 26.5 psia (1.8 bar)
Nominal Stack Temperature = 333 oF (167 oC)
Gas Temperature = 1000 oF (538 oC)
D/A Pressure
Tst - Tst,nominal
Figure 4. Example of the effect of sliding pressure on stack temperature and pressure of a floating integral deaerator
11.1.1.3 Effect of Boiler Pressure on Pinch and Approach
Raising the main pressure reduces steam production rate, and hence rate of heat transfer across the boiler
heat exchange surfaces. This makes the boiler appear oversized relative to its design-point, since the fixed
surface areas are now larger in proportion to heat transfer rates. The heat transfer effectiveness therefore
increases and the pinch difference falls as shown in Figure 5.
The increase in the temperature span between feedwater leaving the D/A and main steam saturation
temperatures raises the approach subcooling, as indicated in Fig. 5.
11-5
 M. A. Elmasri, 1990-2002
Chapter 11: Off-Design Behaviour of HRSG’s
1-P HRB Performance
Effect of Steam Pressure on Pinch & Approach
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Nominal Steam Conditions = 900 psia / 864 F (62 bar / 462 C)
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Gas Temperature = 1000 F (538 C)
Pinch
Approach
Figure 5. Example of the effect of sliding pressure on evaporator pinch and economiser approach subcooling
11.1.2 EFFECT OF FLUE GAS TEMPERATURE
If the HRB delivery pressure were controlled to the design value (900 psia or 62 bar in the example) by
modulating the flow resistance at its exit, the gas flow and temperature would uniquely determine
performance, as described below and in §11.1.3.
11.1.2.1 Effect of Flue Gas Temperature on Steam Production and Steam temperature
Increasing the flue gas temperature increases the effective temperature difference between hot gas and
saturation temperature in the main evaporator. Naturally, this results in greater steam production, as shown
in Figure 6. Greater steam production and flow results in higher water/steam-side heat transfer coefficients,
increasing the overall coefficient of heat transfer. However, that increase is less than proportional to the
increased flow rate. The NTU=(UA/mCp), which is the measure of heat transfer capability per unit of
stream heat capacity falls, reducing the effectiveness of the heat exchangers. The reduced evaporator
effectiveness results in a larger temperature difference at the pinch.
It has already been noted that steam production rate is essentially proportional to the difference between the
initial gas temperature and the gas temperature at the pinch (saturation temperature + pinch difference). Due
to the increased pinch at higher gas temperatures, the increase in steam production is less than directly
proportional to the difference between the initial gas temperature and saturation temperature. The increase
in steam production with hotter gas is further weakened by the rise in superheater exit steam temperature,
which reduces exhaust gas heat available for the evaporator relative to that absorbed by the superheater.
11-6
 M. A. Elmasri, 1990-2002
Chapter 11: Off-Design Behaviour of HRSG’s
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Steam Pressure Controlled to 900 psia (62 bar)
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1-P HRB Performance
Effect of Gas Temperature on Steam Flow & Temperature
F
The increase in final steam temperature does not keep up with increases in gas temperature because of the
higher steam flow and diminishing superheater effectiveness. Thus, in this example, Fig. 6 shows that the
increase in steam temperature is only about 60% of the increase in gas temperature.
-30
-40
-50
100
Steam Flow
Ts - Ts,nominal
Figure 6. Example of the effect of flue gas temperature on steam generation and steam temperature
11.1.2.2 Effect of Flue Gas Temperature on Stack Temperature & Integral D/A Pressure
Fig. 7 shows the effect of initial gas temperature on the cold-end of the boiler. The higher gas temperatures
produce higher steam flows, so the resulting increase in water flow through the main economiser cools the
gas before it reaches the Integral Deaerator/LPB and then the stack. The result is a reduced Integral
Deaerator pressure, if floating, and a reduced stack temperature.
In off-design, the reduction in stack temperature per degree of hot gas temperature rise is less than that
which would be obtained in a new design-point calculation. A new design-point calculation for a higher flue
gas temperature, with the same pinch and approach, would show a need for larger evaporator and
economiser surface areas. Since these areas are already fixed, at off-design the boiler appears undersized
relative to a higher gas temperature.
11.1.2.3 Effect of Flue Gas Temperature on Pinch & Approach
Both pinch difference and approach subcooling increase with gas temperature as shown in Fig. 8, due to the
increased water/steam flow (and heat transfer duty) in relation to the fixed surface area, reducing NTU and
heat exchanger effectiveness.
11-7
 M. A. Elmasri, 1990-2002
Chapter 11: Off-Design Behaviour of HRSG’s
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1-P HRB Performance
Effect of Gas Temperature on D/A & Stack
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Nominal Gas Temperature = 1000 oF (538 oC)
Nominal D/A Pressure = 26.5 psia (1.8 bar)
Nominal Stack Temperature = 333 oF (167 oC)
Steam Pressure Controlled to 900 psia (62 bar)
D/A Pressure
Tst - Tst,nominal
Figure 7. Example of the effect of flue gas temperature on stack temperature and floating IDA pressure
1-P HRB Performance
Effect of Gas Temperature on Pinch & Approach
Tg - Tg,nominal , oC
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Steam Pressure Controlled to 900 psia (62 bar)
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Pinch
Approach
Figure 8. Example of the effect of flue gas temperature on evaporator pinch and economiser approach subcooling
11-8
 M. A. Elmasri, 1990-2002
Chapter 11: Off-Design Behaviour of HRSG’s
11.1.3 EFFECT OF FLUE GAS MASS FLOW RATE
If the HRB delivery pressure were controlled to the design value (900 psia or 62 bar in the example) by
modulating the flow resistance at its exit, the gas flow and temperature would uniquely determine
performance, as described in §11.1.2 and below.
11.1.3.1 Effect of Flue Gas Mass Flow Rate on Steam Production and Steam temperature
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Effect of Gas Flow on Steam Flow & Temperature
F
Increasing the gas flow rate increases steam production. The higher flows improve both water-side and gasside heat transfer coefficients, but at a rate less than linearly proportional to the mass flow rates. Also,
finned-tube heat transfer conductance is reduced by the increased gas-side heat transfer coefficient in
relation to fin conductivity, which reduces fin efficiency and hence heat transfer effective area. Tube wall
and fouling conductances remain constant, and do not increase with mass flow rates at all. Thus, the
increase in overall (UA) is less than the increase in mass flows (mC) and heat exchanger NTU (UA/mC)
falls, reducing effectiveness. Thus, the increase in heat recovery is less than directly proportional to the
increase in gas flow, boiler heat recovery efficiency declines, and stack temperature increases.
-10
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130
Gas Flow, % of nominal
Nominal Steam Temperature = 864 oF (462 oC)
Gas Temperature = 1000 oF (538 oC)
Steam Pressure Controlled to 900 psia (62 bar)
Steam Flow
Ts - Ts,nominal
Figure 9. Example of the effect of flue gas flow rate on steam generation and steam temperature
Figure 9 shows that the final steam temperature drops due to the declining superheater effectiveness. This
causes the increase in steam production rate to be only slightly less than proportional to the increase in flue
gas flow rate, since the less effective superheater allows hotter gas to reach the evaporator. The heat
recovery in the combination (evaporator + superheater) increases at a rate significantly less than proportional
to the increase in flue gas flow rate.
11-9
115
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1-P HRB Performance
Effect of Gas Flow on D/A & Stack
C
 M. A. Elmasri, 1990-2002
Chapter 11: Off-Design Behaviour of HRSG’s
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Gas Flow, % of nominal
Nominal D/A Pressure = 26.5 psia (1.8 bar)
Nominal Stack Temperature = 333 oF (167 oC)
Gas Temperature = 1000 oF (538 oC)
Steam Pressure Controlled to 900 psia (62 bar)
D/A Pressure
Tst - Tst,nominal
Figure 10. Example of the effect of flue gas flow rate on stack temperature and floating IDA pressure
1-P HRB Performance
Effect of Gas Flow on Pinch & Approach
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Pinch or Approach, C
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11-10
Pinch
Approach
 M. A. Elmasri, 1990-2002
Chapter 11: Off-Design Behaviour of HRSG’s
Figure 11. Example of the effect of flue gas flow rate on evaporator pinch and economiser approach subcooling
11.1.3.2 Effect of Flue Gas Mass Flow Rate on Stack Temperature & Integral D/A Pressure
Stack temperature climbs with higher flow rates, as the effectiveness of all heat exchangers declines, as
shown in Figure 10.
The higher gas temperature reaching the Integral Deaerator/LPB increases deaerating steam generation rate.
The higher feedwater flow rates and reduced LTE effectiveness increase demand for deaerating steam.
Depending on which effect is stronger, the pressure of the Integral Deaerator/LPB may float up or down
with changes in flue gas mass flow rate. In our numerical example, the increase in gas temperature does not
increase LP steam production to the point of matching the increased demand for deaerating steam, so the
Integral Deaerator/LPB pressure falls slightly, as shown in Figure 10.
11.1.3.3 Effect of Flue Gas Mass Flow Rate on Pinch & Approach
The decline in evaporator effectiveness increases the pinch, as shown on Figure 11.
The increased pinch means that higher gas temperature reaches the economiser, tending to reduce the
approach subcooling. On the other hand, the reduction in economiser effectiveness tends to increase the
approach subcooling. These competing trends may result in the approach subcooling changing weakly, in
either direction, with increases in flue gas flow rate. In our numerical example the decrease in economiser
effectiveness is the slightly stronger of these competing effects, so approach subcooling falls, ever so
slightly, as seen in Fig. 11.
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Change in H2O, % Volume from nominal of 11%
Nominal Steam Flow = 12.8% of Gas Flow
Nominal Steam Temperature = 864 oF (462 oC)
Nominal Stack Temperature = 333 oF (167 oC)
Gas Temperature = 1000 oF (538 oC)
Steam Pressure Controlled to 900 psia (62 bar)
Stream Flow
Tstack - Tstack,nominal
Tsteam - Tsteam,nominal
Figure 12. Example of the effect of flue gas H2O content on off-design heat recovery
11-11
T - Tnominal
Steam Flow, % of nominal
o
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F
1-P HRB Performance
Effect of Exhaust Water Vapor
C
11.1.4 EFFECT OF FLUE GAS H2O CONTENT
 M. A. Elmasri, 1990-2002
Chapter 11: Off-Design Behaviour of HRSG’s
The combination of massive GT steam injection with natural gas fuel can lead to water vapor content in the
flue gas as high as 22% by volume. This reduces flue gas density and increases its specific heat. The
change in properties raise both HRB draft loss and heat transfer. The overall result is somewhat similar to
increasing gas flow rate: increased steam flow and stack temperature and reduced final steam temperature.
Those effects are illustrated in Fig. 12 for the current HRB example. A 10 % change in H2O content by
volume increases steam production by about 5%, roughly equivalent to the effect of 5% more flue gas.
11.2
SOME OFF-DESIGN CONSIDERATIONS FOR HEAT RECOVERY BOILERS
11.2.1 CONSIDERATIONS FOR SUPERHEATERS & STEAM PIPES:
Superheaters are usually the most critical element in a HRB. The relatively low steam-side heat transfer and
their presence in the hottest gas section results in high metal temperatures. The fact that they carry relatively
low density steam, not water, results in relatively high internal flow velocities and pressure drops. Thus, a
design that is adequate for nominal conditions may operate unsatisfactorily at extreme off-design conditions.
Superheaters are normally designed for steam velocities in the range 25-80 ft/s (8-25 m/s). Designing with a
low velocity results in poor internal heat transfer, which raises metal temperature and necessitates a larger
surface area. Designing with a high velocity increases steam pressure drop and risks excessive off-design
velocities which increases erosion and corrosion. It should also be noted that superheater pressure drop is
"thermodynamically expensive" in a heat recovery boiler compared to economiser pressure drop. A large
superheater pressure drop raises saturation pressure for a given final steam pressure, thereby reducing heat
recovery and steam production; whereas a large economiser pressure drop only slightly increases the
feedpump duty, with a negligibly small thermodynamic impact.
While on this subject, it is worthwhile to note that in conventional steam plants with high temperature
furnaces, superheater pressure drop has a negligible thermodynamic impact, similar to that of economiser
pressure drop. This is because in conventional boilers, heat recovery is not tangibly affected by changes in
saturation temperature, which have a negligible impact on the overall temperature difference between the
radiant flame and the waterwalls in the furnace.
If a heat recovery boiler is operated with nominal flue gas conditions but at reduced steam pressure, the
increased steam production coupled with the decreased density increase the steam velocity and pressure drop
in both the superheater tubes and the steam delivery pipes. This can increase the possibility of erosion,
particularly if the steam has solid impurities such as exfoliated scale. This problem is more likely in process
steam boilers, or in situations where steam flow rate and pressure are decoupled. In typical combined
cycles, operated with sliding pressure, the fall in pressure is proportional to the fall in steam mass flow rate,
resulting in an invariant volumetric flow rate and velocity.
11.2.2 DESUPERHEATING TO CONTROL FINAL STEAM TEMPERATURE:
The variation in final steam temperature as operating conditions change can be problematic to a process or
steam turbine. It is therefore common to design the HRB so that at the operating condition for which steam
temperature is lowest, it would still be adequate. This is usually the combination of highest exhaust gas
mass flow rate and lowest exhaust gas temperature. This implies that at other conditions, when gas flow is
lower and/or hotter, steam temperature will be higher than needed. It would then need to be reduced by
desuperheating (also called attemperation) i.e. mixing a colder stream, water or steam, into the main stream
such that the mixture emerges at the correct temperature. This may be done after the steam leaves the
superheater or between sections of superheater.
11.2.2.1 Effect of Desuperheating Source Temperature
Mixing cold and hot streams is irreversible. Energy is preserved but exergy is destroyed. The exergy
destruction increases with temperature difference between the cold and hot streams. As mentioned in the
11-12
 M. A. Elmasri, 1990-2002
Chapter 11: Off-Design Behaviour of HRSG’s
introduction, there is always a penalty to be paid for destroying exergy. This may be a performance penalty
for a given capital cost or a capital cost penalty to reduce exergy destruction elsewhere and achieve a given
performance.
100.8
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100.6
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100.4
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100.2
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100
-4
99.8
-5
99.6
Stack Temperature
Relative to Nominal, F
Final Steam Flow, %
If the HRB described in the above example were operated with hotter gas, say at 1100 °F (593 °C), but with
steam temperature limited to 850 °F (454 °C). This can be done by desuperheating with water drawn from
makeup (at 59° F/15 °C), or drawn from the deaerator (at 220 °F/104 °C), or drawn from the main
economiser exit (at 485 °F/252 °C). Fig. 13 shows that the final steam flow would be higher and the final
stack temperature lower, if the desuperheating water were drawn from the hottest of those three sources.
Using the hottest water source involves less irreversibility and is therefore more efficient, extracting more
steam from the same hot gas flow and heat transfer surface area.
Steam Flow
Stack Temp.
-6
Desup. Water from
Makeup
D/A 220 F/104C HPE Exit 485F/252C
Desuperheating Source
Figure 13. Example showing how judicious selection of desuperheating water source improves performance
This result may be better understood by noting that use of cold water (from makeup, say) results in less
water being added to the steam to cool it to the required final temperature. By contrast, if hot water from the
main economiser exit is used, a larger amount is needed. This additional water is drawn through both
economisers, LTE and main economiser, increasing their internal heat transfer coefficients and reducing
their exit temperatures, thereby extracting additional heat from the flue gases and lowering the stack
temperature.
Using saturated steam from the evaporator as the attemperating medium produces results similar to those
obtained by using hot water from the economiser exit. The irreversibility of mixing saturated water or
saturated steam into a superheated stream is the same, since they are at the same temperature, although their
enthalpies vary considerably.
Figure 13 shows that a non-negligible performance gain can result from judicious selection of the
desuperheating water source. Unfortunately, most combined cycles and process plants are designed with the
desuperheating water taken from the most obvious source, the main feedpump delivery, largely without
awareness of the thermodynamic penalty involved. Desuperheating water needs to be at a substantial
overpressure relative to the steam into which it is injected, so that its flow rate can be properly modulated by
a control valve.
To use the warmest water source, the main economiser exit, would typically require that the main feedpump
control valve is at the exit of the main economiser, rather than at the exit of the feedpump upstream of the
economiser. This means that the feed pump’s overpressure, typically 25-40%, is transmitted through the
main economiser rather than being throttled in the control valve upstream of the economiser. This may
11-13
 M. A. Elmasri, 1990-2002
Chapter 11: Off-Design Behaviour of HRSG’s
require thicker economiser tubing in some cases, but not in others, since at moderate pressures, tubing
thickness is frequently determined by mechanical integrity needed for finning and bending, as well as by
stiffness to resist vibration, rather than by pressure. Besides, a thicker but commonly used standard tube
may be less expensive than a thinner but infrequently used standard tube. In any case, even if thicker and
costlier economiser tubes and headers are needed, their incremental cost will likely be justified in terms of
the improved thermodynamic performance. Another advantage to placing the feedpump control valve at the
main economiser’s exit is that the overpressure can suppress economiser steaming at off-design. This allows
a lower design point approach subcooling, which provides further performance benefits at design as well as
off-design operation.
In designs where the feedpump control valve is at its exit, upstream of the economiser, the warm economiser
discharge may still be used for desuperheating in conjunction with a small booster pump. The incremental
cost and electricity consumption will likely be more than offset by the superior thermodynamic performance.
11.2.2.2 Choosing Desuperheating Location
Desuperheating to the final temperature after the steam has gone through the entire superheater provides the
simplest design and temperature control system. However, with heavy firing of a duct burner at the HRB's
inlet, high steam and metal temperatures will arise at the superheater's exit. This raises cost by necessitating
more expensive materials.
Carbon steel is the preferred, least expensive, most thermally-conductive material for boiler tubes. It should
be limited to about 900 °F (485 °C) peak metal temperature to avoid excessive corrosion and scaling.
Common practice is to limit it to 750 °F (400 °C) steam temperature. The 400-series stainless steels can be
operated up to about 1200 °F (650 °C) peak metal temperatures, but their conductivity at 900 °F is only
about 65% that of carbon steel. To operate with supplementary firing in the 1650-1950 °F (900-1065 °C)
temperature range, nickel-based alloys can be used, but their cost is significantly higher.
With high temperatures, bare tubes or finned tubes with short fins and lower fin densities are used to keep
peak metal temperatures at acceptable values. This increases cost, since more tubes are needed to provide a
given total heat transfer area, requiring more welds, as well as more headers in designs which employ them.
One way to reduce superheater metal temperatures with duct firing is to desuperheat between two sections of
superheater. The steam goes first through a primary superheater, then water is injected to cool it, after which
it passes through the secondary superheater. The secondary superheater facing the duct burner contains a
higher mass flow rate of cooler steam and operates with lower metal temperature. This practice of
“Interstage Desuperheating” is virtually neutral in its thermodynamic result, i.e. the final steam production
from the same hot gas and heat transfer surface area is almost the same with interstage desuperheating as
with simply desuperheating at the exit. Numerical experiments suggest that interstage desuperheating gives
slightly less final steam with a cool desuperheating water supply (such as makeup) and slightly more final
steam with a hot desuperheating water supply (such as economiser exit).
In the absence of a thermodynamic benefit, the decision to use interstage desuperheating should be based
mostly on cost and complexity of the HRB. In a typical example with supplementary firing to 1200 °F (650
°C), interstage desuperheating can lower maximum tube temperatures by about 80 °F (45 °C), allowing a
less expensive material. However, an adequate length of pipe is needed after the water spray to ensure
complete mixing and evaporation of the water spray. Also, there is the risk of injecting too much water,
carrying over into the secondary superheater, and promoting cracks by thermal fatigue, particularly if the
control system and instrumentation are not designed or maintained properly.
11.2.3 DUCT BURNER LOCATION
Placing the duct burner at the entrance to the HRB simplifies its design and construction. It helps to smooth
distorted exhaust velocity and temperature profiles, without the need for a grid which may otherwise be used
11-14
 M. A. Elmasri, 1990-2002
Chapter 11: Off-Design Behaviour of HRSG’s
and which introduces a parasitic draft loss. Also, this location allows spacing the duct burner away from the
tubes without excessive spacing between HRB tube sections.
Placing the duct burner part way into the HRB, after the exhaust gas has already been cooled by passing
through some superheater sections, lowers peak gas and metal temperatures. In a typical example with
supplementary firing that raises exhaust gas temperature by 250 °F (140 °C), placing the duct burner halfway through the superheater reduces maximum tube temperatures by about 120 °F (70 °C).
There is, however, a minor thermodynamic penalty, on the order of about 0.1% in final steam production,
when the duct burner is placed half-way through the final superheater. This arises from the fact that the
superheater surface upstream of the duct burner in the gas flow does not benefit from the increase in
temperature, wasting some of the duct burner’s fuel temperature potential. The potential of the duct burner
temperature rise to increase heat transfer rate across a given surface by raising the temperature difference
across this surface is thus partially wasted. From a second law perspective, one may note that the mean
temperature of heat addition to the gases is lower, so the irreversibility of combustion/dilution is higher,
causing a thermodynamic loss when the duct burner is placed at a cooler location.
11.3
DUAL-PRESSURE HEAT RECOVERY BOILERS
The off-design behaviour of a dual pressure system is discussed through the numerical example of the
system illustrated in Figs. 14a & 14b.
100 m
15.08 p
1000 T
100 M
71.14 %N2
14 %O2
3 %CO2+SO2
11 %H2O
0.8568 %Ar
59 T
15.6 M
187 T
23.07 p
236 T
12.73 M
2.871 M
0.7965 M
0.7965 M
V2
V4
900 p 835 T 12.6 M
150 p 430 T 2.842 M
LTE
218 T
100 M
LPB
HPE1
IPB
IPS1
HPE3
HPB1
HPS3
1000 T
100 M
IPE2
160.3 p 937.6 p
160.7 p
23.07 p
349 T
432 T 518 T
364 T
236 T
2.842 M 12.73 M
2.842 M
347 T
0.7965 M
296
326
394
489
494
p[psia], T[F], M[lb/s], Steam Properties: IAPWS-IF97
0 10-02-2002 15:51:48 file=C:\SEMINAR\Sem_5-02\Sec11_HRSG_OD\2pHRB_9-02.gtm
23.07 p
187 T
15.6 M
937.6 p
537 T
12.6 M
584
931.4 p
837 T
12.6 M
901
1000
Figure 14a. Nominal design-point heat & mass balance of the dual-pressure HRSG used in the numerical examples
The behaviour of a dual-pressure boiler follows the same rules and principles outlined above for singlepressure systems. The differences arise from the interactions between HP and IP. If the second pressure
(IP) components are all behind the HP, both behave much like single pressure systems, with the gas leaving
11-15
 M. A. Elmasri, 1990-2002
Chapter 11: Off-Design Behaviour of HRSG’s
the HP section playing the role of the initial gas entering the IP system. The influence of the HP on the IP is
strong, and acts through changing that gas temperature. If the entire IP section is behind the HP in the gas
path, it would have no influence on the HP. Normally, however, for efficiency, part of the HP economisers
would be downstream of the IP evaporator in the gas path. Changes in IP operating condition would thus
feed forwards and influence the HP by changing the temperature of the water supplied to the HP
economisers. That effect is usually secondary but not negligible and depends on the economiser
configuration.
LTE
1998.1
1100
LPB
760.6
1000
IPE2
327
HPE1
1447.8
IPB
2493
HRSG Temperature Profile
HPE3
2380.2
IPS1
116.2
HPS3
2766.4
HPB1
8645
0
Q BTU/s
900
5
800
TEMPERATURE [F]
700
7
600
8
10
500
12
400
17
300
14
200
UA BTU/s-F
100
10.26
15.07
0
0
37.41
22.39
5.145
56.07
1.304
23.74
5
10
11.06
15
20
25
HEAT TRANSFER FROM GAS [.001 X BTU/s]
Figure 14b. Nominal design-point temperature profile of the dual-pressure HRSG used in the numerical examples
11.3.1 PRESSURE-FLOW CHARACTERISTICS
11.3.1.1 Effect of Varying the HP Pressure
If the IP in Fig. 14 were maintained at 150 psia (about 10 bar) and the HP pressure varied, such as by
modulation of a valve at its exit, the HP flow and temperature would behave as described for a singlepressure system. The gas temperature reaching the IP would increase with HP pressure, increasing the IP
flow and temperature as shown on Figures 15 & 16. Because the IP compensates for changes in HP heat
absorption, the effect of varying HP on the stack temperature is very weak, as shown on Fig. 16.
Increasing the HP pressure reduces its pinch and increases its approach subcooling, as for a single-pressure
system. The increased IP steam flow reduces effectiveness of IPB and IPE, so both IP pinch and approach
increase. Those effects are of secondary importance.
11-16
 M. A. Elmasri, 1990-2002
Chapter 11: Off-Design Behaviour of HRSG’s
2-P HRB Performance
Effect of HP on Steam Flows
4
13
3
12
2
11
IP Steam Flow, % of Gas Flow
HP Steam Flow, % of Gas Flow
14
1
20
40
60
80
100
120
140
160
180
HP Delivery Pressure, % of nominal
IP Controlled to 150 psia (10.3 bar)
Gas Temperature = 1000 oF (538 oC)
Nominal HP = 900 psia (62 bar)
HP Steam Flow
IP Steam Flow
Figure 15. Typical example showing the effect of varying the HP on both the HP and the IP steam generation rates in a
dual-pressure boiler
2-P HRB Performance
Effect of HP on Temperatures
10
5
0
0
-10
-5
o
Temperature, C
10
o
Temperature, F
20
-10
-20
-15
-30
20
40
60
80
100
120
140
160
180
HP Delivery Pressure, % of nominal
o
o
Gas Temperature = 1000 F (538 C)
Nominal HP Conditions = 900 psia / 838 oF (62 bar / 448 oC)
Nominal IP Conditions = 150 psia / 421 oF (10.3 bar / 216 oC)
Nominal Stack Temperature = 216 oF (102 oC)
THP - THP,nominal
TIP - TIP,nominal
Tstack - Tstack,nominal
Figure 16. Typical example showing the effect of varying the HP on both the HP and the IP superheater exit steam
temperatures and on the stack temperature in a dual-pressure boiler
11-17
 M. A. Elmasri, 1990-2002
Chapter 11: Off-Design Behaviour of HRSG’s
11.3.1.2 Effect of Varying the IP Pressure
2-P HRB Performance
Effect of IP on Steam Flows
4
13
3
12
2
11
IP Steam Flow, % of Gas Flow
HP Steam Flow, % of Gas Flow
14
1
0
50
100
150
200
250
300
IP Delivery Pressure, % of nominal
HP Controlled to 900 psia (62 bar)
Gas Temperature = 1000 oF (538 oC)
Nominal IP = 150 psia (10.3 bar)
HP Steam Flow
IP Steam Flow
Figure 17. Typical example showing the effect of varying the IP on both the HP and the IP steam generation rates in a
dual-pressure boiler
2-P HRB Performance
Effect of IP on Temperatures
80
40
20
20
10
0
0
o
40
-10
-20
Temperature, C
30
o
Temperature, F
60
-20
-40
-30
-60
0
50
100
150
200
250
300
IP Delivery Pressure, % of nominal
o
o
Gas Temperature = 1000 F (538 C)
Nominal HP Conditions = 900 psia / 838 oF (62 bar / 448 oC)
Nominal IP Conditions = 150 psia / 421 oF (10.3 bar / 216 oC)
Nominal Stack Temperature = 216 oF (102 oC)
THP - THP,nominal
TIP - TIP,nominal
Tstack - Tstack,nominal
Figure 18. Typical example showing the effect of varying the HP on both the HP and the IP superheater exit steam
temperatures and on the stack temperature in a dual-pressure boiler
11-18
 M. A. Elmasri, 1990-2002
Chapter 11: Off-Design Behaviour of HRSG’s
If the IP in Fig. 14 were maintained at 900 psia (62 bar) and the IP pressure varied, such as by modulation of
a valve at its exit, the characteristics shown in Figures 17 & 18 are obtained.
Increasing IP pressure reduces its steam production and increases its steam temperature. The IP pressureflow curve differs from that of the HP, in that it is much less "stiff" than for the HP. The percentage change
in steam generation rate relative to the percentage change in pressure is much higher for the IP than for the
HP. For example, Fig. 17 shows that doubling the IP pressure reduces IP steam generation rate by about
30%. By contrast, Fig. 15 shows that doubling the HP pressure reduces HP steam generation rate by only
about 10%. Doubling the HP from its (usually high) nominal design-point value during the operation of a
typical boiler is inconceivable, but raising the IP pressure to a multiple of its (usually modest) nominal
design-point value, may be quite practical. Thus, it is possible in many practical designs to completely shutoff IP flow by "Bottling-up" the drum without draining it, with its pressure rising to an acceptable value,
equal to saturation pressure at the temperature of the exhaust gas reaching its heat exchangers.
Changes in IP pressure have a weak "feed-forward" to the HP performance. Increasing IP pressure
decreases IPB steam generation rate, delivering hotter gas to HPE1 behind it in the gas path. This increases
HP water temperature to the HPE upstream of the IPB in the gas path, and consequently to the HPB, thereby
slightly increasing HP steam generation rate and reducing HP steam temperature as seen in Figs. 17 & 18.
The effect of IP pressure on stack temperature is significant because it is the last heat recovery opportunity
before the stack; as shown in Fig. 18.
The influence of IP pressure on HP pinch is negligible, but HP approach subcooling is reduced with
increasing IP pressure, for the reasons described above. IP pinch decreases as the IP pressure is raised and
the IP flow thus reduced. The IP approach subcooling is under two conflicting effects. Increasing the
pressure reduces flow, increases IPE effectiveness, and tends to reduce subcooling; on the other hand, the
temperature span between feedwater and IP saturation increases with pressure, tending to increase
subcooling. In the numerical example of Fig. 14 the former effect dominates and IP approach subcooling is
reduced with increasing IP pressure.
11.3.2 EFFECT OF EXHAUST GAS TEMPERATURE:
The effects of varying the flue gas initial temperature entering the HRB are felt primarily by the HP, which
behaves much as though it were a single-pressure. Increasing flue gas initial temperature would produce
more HP steam. The increased HP water heating duty upstream of the IP section would therefore extract a
larger temperature drop from the gases before they reach the IP section. Thus, IP steam generation may
increase, but much less percentage wise than HP steam generation, or may actually decrease, or stay
relatively unaffected.
With both HP and IP delivery pressures regulated to their design-point values, the effect of varying flue gas
initial temperature is illustrated by Figures 19-21 and further described below.
Raising the initial gas temperature increases HP steam flow, steam temperature and pinch, much as for the
single-pressure system. The increased water flow through the HPE upstream of the IPB reduces gas
temperature reaching the IPB, in spite of the increased HP pinch. The result, in this example, is a slight
reduction in IP steam flow and temperature as seen in Figs. 19 & 20. In other cases, especially if the IP is
much closer to the HP, IP steam production may increase instead of decreasing, so the general rule is that
effect of initial gas temperature on IP steam production is weak, and could be in either direction. The
economisers between the IPB and the stack carry both the HP and the IP water. The increased total water
flow with increased total steam generation thus reduces stack temperature as seen in Fig. 20.
11-19
 M. A. Elmasri, 1990-2002
Chapter 11: Off-Design Behaviour of HRSG’s
2-P HRB Performance
Effect of Gas Temperature on Steam Flows
o
-25
0
25
50
15
5
14
4
13
3
12
2
11
1
10
-100
IP Steam Flow, % of Gas Flow
HP Steam Flow, % of Gas Flow
-50
Tg - Tg,nominal , C
0
-50
0
50
o
Tg - Tg,nominal , F
HP Controlled to 900 psia (62 bar)
IP Controlled to 150 psia (10.3 bar)
Nominal Gas Temperature = 1000 oF (538 oC)
100
HP Steam Flow
IP Steam Flow
Figure 19. Typical example showing the effect of varying the flue gas initial temperature on both the HP and the IP
steam production in a dual-pressure boiler
2-P HRB Performance
Effect of Gas Temperature on Temperatures
-25
0
25
50
60
30
20
20
10
0
0
-10
-20
-20
-40
-60
-100
o
o
Temperature, F
40
Temperature, C
-50
Tg - Tg,nominal , oC
-30
-50
0
Tg - Tg,nominal , oF
Nominal Gas Temperature = 1000 oF (538 oC)
Nominal HP Conditions = 900 psia / 838 oF (62 bar / 448 oC)
Nominal IP Conditions = 150 psia / 421 oF (10.3 bar / 216 oC)
Nominal Stack Temperature = 216 oF (102 oC)
50
100
THP - THP,nominal
TIP - TIP,nominal
Tstack - Tstack,nominal
Figure 20. Typical example showing the effect of varying the flue gas initial temperature on both the HP and the IP
superheater exit temperatures, and on stack temperature in a dual-pressure boiler
11-20
 M. A. Elmasri, 1990-2002
Chapter 11: Off-Design Behaviour of HRSG’s
Figure 21 shows that at higher gas temperatures, the higher HP flow increases its pinch difference and
approach subcooling as HPB and HPE effectiveness declines. On the other hand, the minor decrease in IP
flow is accompanied by a marginal decrease in IP pinch. IP approach subcooling, however, increases in this
example. This is caused by the increased water flow through HPE1, which is in parallel with the IPE. As
HPE1 becomes cooler with the increased HP water flow, it absorbs more heat from the gas in its path. The
cooler, denser, lower-velocity gas would have a smaller draft loss than that in the parallel IPE path. Since
the draft loss in both paths must be identical, more gas is "siphoned" through the HPE1 side of the parallel
paths, skewing the gas distribution in favor of HPE1 compared to the design point. The decreased gas flow
through the IPE path results in a greater IP approach subcooling.
2-P HRB Performance
Effect of Gas Temperature on Pinch & Approach
-50
-25
Tg - Tg,nominal , oC
0
25
50
60
30
o
40
Temperature, C
o
Temperature, F
50
20
30
20
10
10
0
-100
0
-50
0
o
Tg - Tg,nominal , F
HP Controlled to 900 psia (62 bar)
IP Controlled to 150 psia (10.3 bar)
Nominal Gas Temperature = 1000 oF (538 oC)
50
100
HP Pinch
IP Pinch
HP Approach
IP Approach
Figure 21. Typical example showing the effect of varying the flue gas initial temperature on both the HP and the IP
pinch points and economiser approach subcoolings in a dual-pressure boiler
11.3.3 EFFECT OF EXHAUST GAS FLOW RATE
The effect of changing the flue gas flow rate differs fundamentally from the effect of changing its
temperature. A certain change in gas flow rate propagates equally throughout the boiler, whereas a change
in initial gas temperature impacts the frontmost heat exchangers the most, and is usually accompanied by a
flue gas temperature change in the opposite direction at the rearmost heat exchangers.
Increasing gas flow increases both HP and IP steam generation, as shown in Figure 22. The relative increase
in all mass flows and heat transfer rates reduces the effectiveness of all heat exchangers. The declining
effectiveness of the frontmost heat exchangers in the gas path results in hotter gas reaching the downstream
(in the gas path) heat exchangers. The hotter gas reaching the downstream heat exchangers partly
compensates for their decline in effectiveness, so the downstream heat exchangers receive a proportionately
greater benefit from the increased gas flow than the upstream ones. For instance, in our numerical example
of Fig. 22, increasing the flue gas flow by 30% increases the HP steam production from 12.6% to 16% of the
gas flow, an increase of 27%. On the other hand, it increases the IP steam production from 2.9% to 4% of
the gas flow, an increase of 38%.
11-21
HP Steam Flow, % of nominal Gas Flow
2-P HRB Performance
Effect of Gas Flow on Steam Flows
16
8
14
6
12
4
10
2
8
IP Steam Flow, % of nominal Gas Flow
 M. A. Elmasri, 1990-2002
Chapter 11: Off-Design Behaviour of HRSG’s
0
70
80
90
100
110
120
130
Gas Flow, % of nominal
HP Controlled to 900 psia (62 bar)
IP Controlled to 150 psia (10.3 bar)
o
o
Gas Temperature = 1000 F (538 C)
HP Steam Flow
IP Steam Flow
Figure 22. Typical example showing the effect of varying the flue gas mass flow rate on both the HP and the IP steam
production in a dual-pressure boiler
2-P HRB Performance
Effect of Gas Flow on Temperatures
30
15
o
Temperature, C
10
o
Temperature, F
20
10
5
0
0
-10
-5
-10
-20
70
80
90
100
110
120
130
Gas Flow, % of nominal
o
o
Gas Temperature = 1000 F (538 C)
Nominal HP Conditions = 900 psia / 838 oF (62 bar / 448 oC)
Nominal IP Conditions = 150 psia / 421 oF (10.3 bar / 216 oC)
Nominal Stack Temperature = 216 oF (102 oC)
THP - THP,nominal
TIP - TIP,nominal
Tstack - Tstack,nominal
Figure 23. Typical example showing the effect of varying the flue gas mass flow rate on both the HP and the IP
superheater exit temperatures, and on stack temperature in a dual-pressure boiler
11-22
 M. A. Elmasri, 1990-2002
Chapter 11: Off-Design Behaviour of HRSG’s
HP steam temperature falls with increasing gas flow, just as for a single-pressure boiler, as shown in Figure
23. IP steam temperature is subject to conflicting effects. The reduced IPS effectiveness tends to depress it,
whilst the higher gas temperature reaching it tends to raise it. It may thus change in either direction or
experience no significant variation. It increases very slightly with gas flow in the present numerical
example, as seen in Fig. 23.
Stack temperature rises as heat transfer effectiveness falls.
Figure 24 shows increasing pinch differences as HPB and IPB effectivenesses decline with higher flow and
heat transfer rates. Approach subcoolings are under the conflicting effects of reduced economiser
effectivenesses, tending to increase them, and higher gas temperatures reaching the economisers, tending to
decrease them. In the present numerical example HP approach subcooling decreases, and IP approach
increases, with increased gas flow rate, each by a just few degrees.
2-P HRB Performance
Effect of Gas Flow on Pinch & Approach
70
35
60
o
Temperature, C
o
Temperature, F
30
50
25
40
20
30
15
20
10
10
70
80
90
100
110
Gas Flow, % of nominal
HP Controlled to 900 psia (62 bar)
IP Controlled to 150 psia (10.3 bar)
Nominal Gas Temperature = 1000 oF (538 oC)
120
130
HP Pinch
IP Pinch
HP Approach
IP Approach
Figure 24. Typical example showing the effect of varying the flue gas mass flow rate on both the HP and the IP pinch
points and economiser approach subcoolings in a dual-pressure boiler
11-23
 M. A. Elmasri, 1990-2008
Chapter 12: Triple-Pressure & Reheat CC
TRIPLE-PRESSURE & REHEAT COMBINED CYCLES: DESIGNPOINT THERMODYNAMICS
Content Revised November, 2002. Updated September, 2008
© Maher Elmasri 1990-2008
12.1
INTRODUCTION
The dual-pressure combined cycle improves performance by increasing heat recovery, generating additional
steam at a lower pressure for induction into the steam turbine. An obvious extension is to increase the
number of pressures to three (or more), further increasing heat recovery and lowering stack temperature.
This approach is not the most effective, thermodynamically or economically. Even if the resulting low stack
temperature were tolerable, there is a diminishing return. The additional low-temperature heat recovered
has low exergy, and so its contribution to power output is modest. It requires increasing the boiler heat
transfer surface area, steam piping, condenser surface area, cooling water flow, etc, disproportionately to the
improved performance, so the additional cost and complexity reduce, or may even nullify, the value of the
efficiency and output gains.
To illustrate the principles, we consider different cycle types, all built around the same gas turbine model, a
180-MW-class 60-Hz machine (GE Frame 7FB). For each cycle type, the key design parameters, such as
steam pressures, are selected appropriately for the configuration. The general thermodynamic design
assumptions, such as pinch points, are the same for all types, and are quite similar to those for the base cases
in Chapters 9 and 10. For consistency, wherever possible, design assumptions are the same for all cycle
types, even at the expense of some realism. For example, all cycles are assumed to be designed with a boiler
draft loss of 26 mb (10.4" H2O), even though a single pressure boiler would typically be designed with a
slightly smaller draft loss. Likewise, steam temperature at the throttle is assumed at 565 °C (1049 °F) for all
cycles, even though this value is more typical of reheat than non-reheat cycles. HP and IP pinch differences
are both assumed at 13 °C (23.4 °F) and, for triple-pressure cases, LP pinch is 10°C (18 °F). All plants are
designed for ISO conditions, with a water-cooled condenser and a mechanical draft cooling. Condenser
pressure is assumed at 50 mb (0.725 psia) for all cycles, even though practical reheat cycles may be
designed at ISO conditions with a slightly lower condenser pressure, and non-reheat cycles may be designed
with a slightly higher condenser pressure, to ensure that steam turbine exhaust quality is above
approximately 88%. Thus, the only significant differences between the various cycle designs considered are
the basic configuration, and the steam pressures commensurate with it.
12.2
NON-REHEAT CYCLES WITH 1, 2 AND 3 PRESSURES
HP throttle steam conditions are selected at 80 bar/565 °C (1160 psia/1049 °F). The condenser pressure of
50 mb (0.725 psia) results in about 87.5% steam quality at the exhaust, before leaving losses, which
corresponds to about 88.5% quality at the condenser. This is at the low end of the allowable range for steam
turbines without moisture removal.
12.2.1 SINGLE-PRESSURE
Figures 1a & 1b show this design. Net plant output comes to 265.6 MW at 53.64 % net efficiency. The
required boiler heat transfer surface area is calculated as 88,800 m2. (956,000 ft2). For a basic boiler,
without by-pass stack, supplementary firing, SCR, or CO catalyst, the estimated dry weight, including stack,
is ~1080 tonnes, and estimated US cost is ~$12 million.
12-1
 M. A. Elmasri, 1990-2008
Chapter 12: Triple-Pressure & Reheat CC
As would be expected for a single-pressure HRB, the stack temperature is high, at 133 °C (271 °F) and heat
recovery efficiency is only 81.3 %. The heat recovered is of high-grade, going into high-pressure steam.
Bottoming cycle internal efficiency, power per unit of heat recovered, is therefore high at 36.44 %.
12.2.2 DUAL-PRESSURE
Figures 2a & 2b show this case, with the HP at 80 bars (1160 psia) and the IP at 8 bars (116 psia).
Compared to the single-pressure example, the net performance improves by 5 MW to 270.5 MW (a gain of
1.8%), with a net plant efficiency of 54.62 %. Figures 4 through 11 show the differences between the dualpressure cycle and the base, single-pressure design.
Figure 4 shows that the second pressure provides a significant increase in heat recovery, lowering stack
temperature by 46 °C (83 °F). This additional heat recovery generates steam at the lower pressure of 8 bars,
less effective in producing power than the main 80 bar steam. After this additional steam has done its
smaller share of turbine work, it goes to the condenser, and Figure 5 shows that the steam mass flow rate to
the condenser is 11% greater than in the single-pressure case. Thus, as seen in Figure 6, although a
significant reduction in stack energy loss (7% of exhaust sensible heat) is accomplished by introducing the
second pressure, three-quarters of the additional heat thus recovered is dumped into the condenser.
With the additional heat recovered generating less effective steam, the increase in power output is less than
proportional to the increase in heat recovered. Thus, the dual-pressure cycle has a lower internal efficiency
(35.48%) than the single-pressure design (36.44%), as seen in Figure 7. The substantial increase in heat
recovery efficiency, from 81.3% to 88.3%, as seen in Figure 8, is greater than the reduction in steam cycle
internal efficiency, so the bottoming cycle overall efficiency improves, from 29.62% to 31.33%, as shown in
Figure 9.
The additional heat recovery, along with the narrower temperature differences along the boiler temperature
profile (compare Figures 1b and 2b) increase the required heat transfer surface area, by 56% over that for the
single-pressure base case design, as shown in Figure 10. This increases basic ∗ boiler estimated dry weight,
including stack, to ~1470 tonnes and its estimated US cost to ~ $15 million. Additionally, the dual-pressure
plant requires a larger steam turbine generator, larger steam turbine expansion path at its low-pressure end,
as well as a larger condenser and associated cooling equipment.
Overall cost of the dual-pressure plant is estimated to be 6% higher than the overall cost of the singlepressure plant, but since it produces 1.8% more power, its specific cost in $/kW is only 4.2% higher. Its
incremental capacity cost is about $2000 per additional kW, about triple the cost per kW of the base singlepressure combined cycle. On a pure cost per installed kW of capacity, it cannot be justified, but the
improvement in heat rate over the life of the plant will, in most cases, justify the additional cost. This is
discussed more fully in the chapter on "Cost-Efficiency Tradeoffs". Another way of looking at the
difference is to realise that the incremental capacity is obtained without burning additional fuel. Since a
fuel-free kW of capacity is usually worth roughly three times as much as a fuel-burning kW, gaining fuelfree kilowatts at an incremental cost that is triple the capital cost of the fuel-burning kilowatts is usually well
worthwhile.
All boiler figures in this chapter are for a “basic HRSG” without by-pass damper, by-pass stack, supplementary firing,
SCR, or CO catalyst. Price indicated is equipment only, excluding foundations, feedpumps, erection, start-up, etc.
∗
12-2
 M. A. Elmasri, 1990-2008
Chapter 12: Triple-Pressure & Reheat CC
GT PRO 10.9 Maher Elmasri
Net Power 265651 kW
LHV Heat Rate 6712 kJ/kWh
1.01 p
15 T
60 %RH
440.4 m
0 m elev.
450.3 m
1X GE 7251FB
18.55 p
416 T
17.62 p
1369 T
74.29 %N2
12.16 %O2
3.914 %CO2+SO2
8.742 %H2O
0.8947 %Ar
1.04 p
632 T
450.3 M
181886 kW
1p
15 T
440.4 m
91743 kW
CH4 9.896 m
LHV= 495277 kWth
25 T
33 T
74.43 M
95 T
80 p
565 T
73.68 M
87 T
0.05 p
33 T
73.66 M
1.2 p
105 T
74.42 M
1.371 M
1.385 M
0.0144 M
81.6 p 567 T
33 T
FW
LTE
LPB
HPE2
HPE3
1.2 p
95 T
74.43 M
1.2 p
105 T
1.371 M
85.72 p
269 T
74.42 M
84.46 p
296 T
74.42 M
HPS0
HPB1
HPS3
133 T
450.3 M
631 T
450.3 M
173
180
p[bar], T[C], M[kg/s], Steam Properties: Thermoflow - STQUIK
180
180
291
84.46 p 83.03 p
299 T 310 T
73.68 M 73.68 M
312
519
GT PRO10.9 Maher Elmasri
LTE
19372
700
Net Power 265651 kW
LHV Heat Rate 6712 kJ/kWh
HRSGTemperature Profile
HPB1
105451
HPE3
10423
LPB
3077
HPE2
54167
81.6 p
567 T
73.68 M
529
631
HPS0
4950
HPS3
54352
0
Q kW
600
5
4
TEMPERATURE [C]
500
400
9
300
200
7
1714
UA kW/C
100
219.7
0
0
1299.8
22.67
565.4
43.23
50
1446.5
100
150
200
431.8
250
300
HEAT TRANSFER FROMGAS [.001 X kW]
Figures 1a (top) and 1b (bottom). Flow schematic and heat recovery temperature profile of the single-pressure, nonreheat combined cycle used in the examples
12-3
 M. A. Elmasri, 1990-2008
Chapter 12: Triple-Pressure & Reheat CC
GT PRO 10.9 Maher Elmasri
Net Power 270529 kW
LHV Heat Rate 6591 kJ/kWh
1.01 p
15 T
60 %RH
440.4 m
0 m elev.
450.3 m
1X GE 7251FB
18.55 p
416 T
17.62 p
1369 T
74.29 %N2
12.16 %O2
3.914 %CO2+SO2
8.742 %H2O
0.8947 %Ar
1.04 p
632 T
450.3 M
181886 kW
1p
15 T
440.4 m
97045 kW
CH4 9.896 m
LHV= 495277 kWth
25 T
33 T
81.73 M
95 T
80 p
565 T
73.68 M
87 T
0.05 p
33 T
81.69 M
1.2 p
105 T
0.0144 M
81.6 p 567 T
8 p 258 T
81.79 M
1.506 M
1.441 M
8.031 M
33 T
FW
LTE
LPB
IPE2
1.2 p
95 T
81.73 M
1.2 p
105 T
1.506 M
8.653 p
171 T
81.79 M
IPB
HPE2
HPE3
IPS2
HPB1
HPS0
HPS3
631 T
450.3 M
87 T
450.3 M
132
139
p[bar], T[C], M[kg/s], Steam Properties: Thermoflow - STQUIK
8.653 p 85.72 p
174 T 269 T
8.031 M 74.42 M
187
220
84.46 p 8.32 p 84.46 p 83.03 p
260 T 299 T 310 T
296 T
74.42 M 8.031 M 73.68 M 73.68 M
287
309
312
519
81.6 p
567 T
73.68 M
529
631
GT PRO10.9 Maher Elmasri
LTE
21210
700
LPB
3355
IPB
15995
IPE2
23003
HRSGTemperature Profile
HPE2
33135
IPS2
1609.9
HPE3
10423
Net Power 270529 kW
LHV Heat Rate 6591 kJ/kWh
HPB1
105451
HPS0
4950
HPS3
54352
0
Q kW
600
5
4
TEMPERATURE [C]
500
400
76
9
300
10
12
200
1714
UA kW/C
100
474.3
0
0
972.7
111.1
615.2
50
1105.1
687.2
18.64
100
22.67
1446.5
150
200
431.8
250
300
HEAT TRANSFER FROMGAS [.001 X kW]
Figures 2a (top) and 2b (bottom). Flow schematic and heat recovery temperature profile of the dual-pressure, nonreheat combined cycle used in the examples
12-4
 M. A. Elmasri, 1990-2008
Chapter 12: Triple-Pressure & Reheat CC
GT PRO 10.9 Maher Elmasri
Net Power 271440 kW
LHV Heat Rate 6569 kJ/kWh
1.01 p
15 T
60 %RH
440.4 m
0 m elev.
450.3 m
1X GE 7251FB
18.55 p
416 T
17.62 p
1369 T
74.29 %N2
12.16 %O2
3.914 %CO2+SO2
8.742 %H2O
0.8947 %Ar
1.04 p
632 T
450.3 M
181886 kW
1p
15 T
440.4 m
98103 kW
CH4 9.896 m
LHV= 495277 kWth
25 T
33 T
83.16 M
95 T
80 p
565 T
73.68 M
87 T
0.05 p
33 T
83.06 M
1.2 p
105 T
81.6 p 567 T
2.646 p 153 T
17 p 295 T
79.61 M
1.504 M
5.054 M
3.558 M 0.0229 M
5.865 M
33 T
FW
LTE
LPE
80 T
450.3 M
LPB
LPS
IPE2
IPB
HPE2
HPE3
IPS2
HPB1
HPS0
HPS3
631 T
450.3 M
IPE1
1.2 p
95 T
83.16 M
3p
2.857 p 18.39 p
155 T
205 T
130 T 134 T
134 T 5.063 M 3.558 M 79.61 M
125
146
169
170
p[bar], T[C], M[kg/s], Steam Properties: Thermoflow - STQUIK
18.39 p 85.72 p
208 T 269 T
5.866 M 74.42 M
221
244
84.46 p 17.68 p 84.46 p 83.03 p
297 T 299 T 310 T
296 T
74.42 M 5.865 M 73.68 M 73.68 M
288
309
312
519
81.6 p
567 T
73.68 M
529
631
GT PRO10.9 Maher Elmasri
LTE
21584
700
LPS
162.1
LPB
11008
IPE2
LPE
24973
537.3
IPE1
9463
HRSGTemperature Profile
HPE2
21790
IPS2
IPB
1341.3
10915
HPE3
10423
Net Power 271440 kW
LHV Heat Rate 6569 kJ/kWh
HPB1
105451
HPS0
4950
HPS3
54352
0
Q kW
600
5
4
TEMPERATURE [C]
500
400
76
9
300
12
200
17
15
10
13
14
UA kW/C
100
570.6
0
0
1017.5
29.62
497.8
6.834
593.4
50
490.4
830.7
663.3
29.94
100
22.67
1446.5
150
200
431.8
250
300
HEAT TRANSFER FROMGAS [.001 X kW]
Figures 3a (top) and 3b (bottom). Flow schematic and heat recovery temperature profile of the triple-pressure, nonreheat combined cycle used in the examples
12-5
 M. A. Elmasri, 1990-2008
Chapter 12: Triple-Pressure & Reheat CC
Figure 4. Stack temperatures for the various bottoming cycle types considered
Exhaust Steam Mass Flow Rate to Condenser, Relative to 1P Cycle
Relative ST exhaust mass flow rate
115
110
105
100
95
90
85
80
1P
2P
3P
1PRH
2PRH
3PRH
2P-HPRH
2P-PMRH
Cycle Type
Figure 5. Relative condensing steam mass flow rates for the various bottoming cycle types considered
12.2.3 TRIPLE-PRESSURE, NON-REHEAT
Figs. 3a & 3b show this case. Figure 4 shows that the third pressures reduces stack temperature by a mere 7
°C (13 °F) compared to the two-pressure. This is due to the high exhaust gas temperature of the gas turbine
used in this series of examples, and the third pressure would have a more significant impact if the gas
turbine had a more modest exhaust temperature. Adding the third pressure continues the trends discussed in
§12.2.2, and displayed in Figures 4 through 10, but the resulting performance gains are more modest, and
the necessary increases in equipment sizes are larger relative to the benefits.
12-6
 M. A. Elmasri, 1990-2008
Chapter 12: Triple-Pressure & Reheat CC
Figure 6. Stack & condenser energy losses for the various bottoming cycles (relative to the single-pressure base case)
Figure 7. Bottoming cycle Internal Efficiencies for the various bottoming cycles considered
Examination of Figures 4 through 10 shows the diminishing return of increasing heat recovery by adding
evaporators at lower pressure. Greater recovery of low-grade, low-exergy heat contributes weakly to power
output, but increases costs disproportionately. On the boiler side, not only is there additional heat recovered,
but this is across a small temperature difference, as seen by comparing Figures 1b, 2b, and 3b. This requires
substantial increases in heat transfer surface area, as seen in Figure 10. Furthermore, lower stack
temperatures may require taller stacks for plume dispersion, and increase the need for stack insulation to
avoid condensation. On the heat rejection side, a larger cooling system is needed to dispose of a greater
proportion of the additional heat recovered. This increases the costs, auxiliary power loads, and the
environmental footprint of the cooling system. If the plant is cooled by a cooling tower, more water will be
consumed, and more power expended on the fans and pumps.
12-7
 M. A. Elmasri, 1990-2008
Chapter 12: Triple-Pressure & Reheat CC
Figure 8. Heat recovery efficiencies for the various bottoming cycles considered
Figure 9. Bottoming cycle Overall Efficiencies for the various bottoming cycles considered
Improving performance through increasing low grade heat recovery also increases steam turbine cost, since
the additional low-pressure steam requires a greater flow passing capacity in the low pressure turbine stages.
These stages, particularly the last few, are relatively expensive because of the long buckets associated with
low-pressure, high-volume steam.
Thus adding pressures to increase heat recovery and lower stack temperature provides diminishing returns in
output, and steep increases in cost. Stack heat rejection is cheap and condenser heat rejection is expensive.
12-8
 M. A. Elmasri, 1990-2008
Chapter 12: Triple-Pressure & Reheat CC
Figure 10. Boiler heat transfer surface areas & estimated costs (relative to the single-pressure base case)
Figure 11. Net combined cycle efficiency for the various cycles considered
12-9
 M. A. Elmasri, 1990-2008
Chapter 12: Triple-Pressure & Reheat CC
12.3
REHEAT CYCLES
A perfect bottoming cycle would have a working fluid with a very low latent heat of vaporization compared
to the sensible enthalpy rise of the vapour. This would allow a close match between the near-linear
temperature profile of cooling the exhaust gases and heating the bottoming cycle working fluid, approaching
an ideal, balanced counterflow heat exchanger for the boiler. Supercritical steam provides a better working
fluid profile, without latent heat, but with a curved enthalpy-temperature profile reminiscent of a low latent
heat. High subcritical pressures provide similar advantages, so do exotic working fluids, such as ammoniawater mixtures.
Reheat is another way to diminish latent heat transfer at constant temperature, and to increase sensible
enthalpy rise at a temperature profile closer to parallel to that of the exhaust gas. It improves bottoming
cycle performance by reducing steam production! Some of the high-grade heat is diverted from additional
steam generation, and used instead to reheat the already generated steam, after its partial expansion in the
steam turbine has cooled it, and before sending it back to the steam turbine for further expansion.
Various configurations of reheat cycles are discussed below, some commonplace and some less
conventional arrangements. In this set of examples, throttle pressure is set at 115 bars (1668 psia). Hot
reheat pressure is set at around 30 bars (435 psia), with minor variations between cycle types, and IP and LP
pressures are selected according to cycle type, as explained below.
12.3.1 SINGLE-PRESSURE REHEAT CYCLE
This configuration is not recommended, unless supplementary firing is used to depress stack temperature, or
for special applications requiring very high stack temperatures. It is included here for completeness, and its
simplicity helps to illustrate the basic effects of reheat.
Figures 12a and 12b illustrate this configuration. The numbers above the temperature profile in Fig. 12b
show the heat transfer rate in each heat exchanger. In this example, about 51 MW of thermal energy goes to
the superheater, and about 26 MW to the reheater, both shown as being in parallel with each other, at the
front of the boiler. In practice, these heat exchangers are likely to be installed in an interleaved, not parallel,
arrangement, but this does not alter the thermodynamics, as discussed in §12.3.4, and the parallel
arrangement is used to help simplify the diagrams and clarify the underlying thermodynamic principles.
The evaporator in this example absorbs about 74 MW. Had there not been a reheater, the 26 MW it takes
would have been available for use by the evaporator and superheater instead, in the ratio 74:51, so the
evaporator would have taken an additional 26*74/(74+51) ≅ 15 MW, and would thus have generated ≈ 21%
more steam. The economisers and deaerator below the pinch absorb about 81 MW in this example. Had
there been no reheater, the additional 21% of steam generated would have required an additional 21% of
feedwater flow, and the heat exchangers below the pinch would have needed ≈ 1.21*81 ≈ 98 MW. The
temperature drop experienced by the flue gases from the pinch to the stack can be found from the
temperatures appearing at the bottom of the boiler duct mimic in Figure 12a. This is 338-171=167 °C in the
present example. Had there been no reheater, this temperature drop would have been roughly 21% higher,
i.e. about 202 °C, so the stack would have been at 338-202 ≈ 136 °C instead of 171 °C. Thus, the presence
of the reheater in this example results in a stack temperature roughly 35 °C (63 °F) higher than it would have
been without reheat.
Comparing Figs. 1a and 12a shows that the stack for the 1PRH cycle is 38 °C higher than the 1P (non-reheat
cycle). Of this difference, roughly 35 °C are due to the reheat, as discussed above, and the remaining 3 °C
due to the higher steam pressure assumed in the 1PRH case.
12-10
 M. A. Elmasri, 1990-2008
Chapter 12: Triple-Pressure & Reheat CC
GT PRO 10.9 Maher Elmasri
Net Power 266079 kW
LHV Heat Rate 6701 kJ/kWh
1.01 p
15 T
60 %RH
440.4 m
0 m elev.
450.3 m
1X GE 7251FB
18.55 p
416 T
17.62 p
1369 T
74.29 %N2
12.16 %O2
3.914 %CO2+SO2
8.742 %H2O
0.8947 %Ar
1.04 p
632 T
450.3 M
181886 kW
1p
15 T
440.4 m
92196 kW
CH4 9.896 m
LHV= 495277 kWth
25 T
33 T
61.99 M
100 T
115 p
565 T
61.37 M
87 T
0.05 p
33 T
60.88 M
1.2 p
105 T
61.98 M
27.6 p 565 T
31.64 p 376 T
116.7 p 567 T
0.5603 M
0.5659 M
0.457 M
59.58 M
59.58 M
33 T
FW
LTE
LPB
HPE3
HPB1
1.2 p
100 T
61.99 M
1.2 p
105 T
0.5603 M
120.8 p
322 T
61.98 M
120.8 p
325 T
61.37 M
HPS3
171 T
450.3 M
631 T
450.3 M
RH3
208
210
210
p[bar], T[C], M[kg/s], Steam Properties: Thermoflow - STQUIK
210
567 T
567 T
484
631
338
GT PRO10.9 Maher Elmasri
LTE
17439
700
LPB
1257.2
Net Power 266079 kW
LHV Heat Rate 6701 kJ/kWh
HRSGTemperature Profile
HPB1
74281
HPE3
62947
RH3
25908
HPS3
51420
500
TEMPERATURE [C]
0
Q kW
600
5
400
7
300
14
17
200
UA kW/C
100
12.15
143.2
0
0
1359.4
492
303.4
1279.1
50
100
150
200
250
HEAT TRANSFER FROMGAS [.001 X kW]
Figures 12a (top) and 12b (bottom). Flow schematic and heat recovery temperature profile of the single-pressure,
reheat combined cycle used in the examples
12-11
 M. A. Elmasri, 1990-2008
Chapter 12: Triple-Pressure & Reheat CC
Because of the high pressure, and because the reheater absorbs heat from the gas above the pinch, the ratio
of heat needed above the pinch to heat needed below the pinch is high. In this example, roughly 151 MW
(74+51+26) are needed above the pinch, and roughly 81 below it, so if supplementary firing is used to raise
the exhaust gas temperature by 100 °, the stack could become about 55 ° cooler (100*81/151). Thus, duct
burner fuel utilisation efficiency would be high, at about 155%.
Comparing the performance of the 1P and the 1PRH in Figures 4 through 11 shows that with a singlepressure cycle, without supplementary firing, reheat provides only a very modest net performance
advantage. It results in a high stack temperature (Figure 4), but the heat that is recovered is of high grade,
and is transferred across smaller average temperature differences on the hot-end of the temperature profile,
as seen by comparing Figs. 1b and 12b. Thus, the recovered heat is converted to power at a very high
internal efficiency (Figure 7). The product of high bottoming cycle internal efficiency with poor heat
recovery efficiency (Figure 8), results in a bottoming cycle overall efficiency (Figure 9) that is only slightly
higher than 1P. Comparing 1P with 1PRH in Figure 6 shows that the reheat increases the stack energy loss,
but reduces the condenser energy loss by a slightly greater extent, resulting in the modest net gain.
12.3.2 DUAL-PRESSURE REHEAT CYCLE
The high stack temperature associated with 1PRH makes the notion of a second pressure more attractive
than without reheat, so the full advantage of reheat can only be realised by installing an IP. Figures 13a &
13b show an example of a dual-pressure reheat cycle. The steam leaving the HPT is combined with the IP
steam and the mixture reheated to the initial temperature before returning to the LPT.
As discussed in §12.3.1, reheat enables efficient utilisation of exhaust energy by concentrating it into a
relatively low mass flow of steam, since the HPT massflow is heated twice in the hot-end of the boiler, less
exhaust energy is available for evaporating steam. The condenser loss is therefore reduced. Comparing 2P
with 2PRH in Figure 5 shows that the mass flow to the condenser drops with reheat compared to non-reheat.
From a second law viewpoint, reheat reduces heat transfer exergy losses. Comparison of Figures 2b and 13b
shows how the latter provides a temperature profile with a more uniform working fluid profile, lying closer
to the exhaust gas heat-source curve, particularly towards the high-exergy, hot end of the profiles. The loss
of exergy associated with temperature degradation is therefore reduced.
It is instructive to compare the dual-pressure reheat (2PRH) and the triple-pressure non-reheat (3P)
examples. The 2PRH case has a much higher stack temperature (Figure 4), corresponding to a significantly
lower heat recovery efficiency (Figure 8). Figure 7 shows, however, that the internal efficiency of utilising
the recovered heat is much higher for 2PRH than for 3P. On average, the kilowatts produced at the steam
turbine per kJ (or BTU) of heat recovered for 2PRH is almost 12% greater than for 3P non-reheat. Figure 6
shows that whereas the 2PRH loses about 7.5% more exhaust energy at the stack, compared to 3P, it rejects
about 8% less exhaust energy at the condenser. The result is a net power gain of about 0.5% of the exhaust
energy. This corresponds to about 0.3 percentage points in net combined cycle efficiency (Figure 11), since
the exhaust energy contains ~ 60% of the gas turbine’s fuel input energy.
The temperature profile of Fig. 13b is greatly superior to that of Fig. 3b. The temperature differences are
more uniform overall. Smaller temperature differences are employed at the hot-end of the profile, reducing
exergy loss in the valuable, high-temperature, high-exergy region. Larger temperature differences prevail in
the cold end, where the heat possesses low exergy, avoiding small temperature differences and large
surfaces that recover low-value heat. Figure 10 shows the 2PRH boiler requires 22% less surface area than
the 3P boiler. The 2PRH boiler, however, is at higher pressure and contains a much greater percentage of
high-temperature surface in superheaters and reheaters, which require costly materials, so the cost advantage
of the 2PRH boilers is small, within the uncertainty inherent in the cost estimating calculation.
12-12
 M. A. Elmasri, 1990-2008
Chapter 12: Triple-Pressure & Reheat CC
GT PRO 10.9 Maher Elmasri
Net Power 272461 kW
LHV Heat Rate 6544 kJ/kWh
1.01 p
15 T
60 %RH
440.4 m
0 m elev.
450.3 m
1X GE 7251FB
18.55 p
416 T
17.62 p
1369 T
74.29 %N2
12.16 %O2
3.914 %CO2+SO2
8.742 %H2O
0.8947 %Ar
1.04 p
632 T
450.3 M
181886 kW
1p
15 T
440.4 m
99134 kW
CH4 9.896 m
LHV= 495277 kWth
25 T
33 T
67.81 M
100 T
115 p
565 T
59.5 M
87 T
0.05 p
33 T
67.32 M
1.2 p
105 T
67.89 M
0.613 M
0.5361 M
0.3927 M
27.6 p 565 T
31.64 p 376 T
116.7 p 567 T
66.05 M
57.75 M
33 T
FW
LTE
LPB
1.2 p
100 T
67.81 M
1.2 p
105 T
0.613 M
IPE2
IPB
HPE3
IPS2
HPB1
HPS3
129 T
450.3 M
631 T
450.3 M
RH3
31.79 p 31.79 p
234 T 237 T
67.89 M 8.303 M
168
171
250
p[bar], T[C], M[kg/s], Steam Properties: Thermoflow - STQUIK
120.8 p 30.57 p 120.8 p
320 T 325 T
322 T
60.1 M 8.302 M 59.5 M
280
334
338
GT PRO10.9 Maher Elmasri
LTE
19031
700
Net Power 272461 kW
LHV Heat Rate 6544 kJ/kWh
HRSGTemperature Profile
HPB1
IPB
72024
14590
IPS2
2005.6
HPE3
26908
IPE2
38235
LPB
1326.8
RH3
29738
HPS3
49858
0
Q kW
600
500
TEMPERATURE [C]
567 T
567 T
480
631
5
400
76
300
11
12
200
14
17
UA kW/C
100
20.5
235.2
0
0
1128.1
1096.7
586.6
50
42.81
100
484.7
344.3
1264.2
150
200
250
300
HEAT TRANSFER FROMGAS [.001 X kW]
Figures 13a (top) and 13b (bottom). Flow schematic and heat recovery temperature profile of the dual-pressure, reheat
combined cycle used in the examples
12-13
 M. A. Elmasri, 1990-2008
Chapter 12: Triple-Pressure & Reheat CC
12.3.3 TRIPLE-PRESSURE REHEAT CYCLE
The 2PRH cycle has a stack 41 °C (74 °F) warmer than 2P (non-reheat). Although we found that going
from dual-pressure to triple-pressure without reheats was of questionable efficacy, it should be worthwhile
with reheat, due to the warm 2PRH stack.
Figures 14a & 14b show an example of a triple-pressure reheat cycle. The third pressure allows reduction of
stack temperature by 38 °C (68 °F) from 2PRH as seen in Figure 4. As one would expect, internal
efficiency drops as seen in Figure 7, due to the additional low-grade heat recovery. Stack loss is reduced but
condenser loss increased by comparison with 2PRH as seen in Figure 6. However, the reduction in stack
loss outweighs the increase in condenser loss, and net combined cycle efficiency increases by 0.75
percentage points as seen in Figure 11. Naturally, cost also increases, with a larger steam turbine to
accommodate the larger exhaust flow, as well as a larger boiler to generate the extra LP steam.
12-14
 M. A. Elmasri, 1990-2008
Chapter 12: Triple-Pressure & Reheat CC
GT PRO 10.9 Maher Elmasri
Net Power 276262 kW
LHV Heat Rate 6454 kJ/kWh
1.01 p
15 T
60 %RH
440.4 m
0 m elev.
450.3 m
1X GE 7251FB
18.55 p
416 T
17.62 p
1369 T
74.29 %N2
12.16 %O2
3.914 %CO2+SO2
8.742 %H2O
0.8947 %Ar
1.04 p
632 T
450.3 M
181886 kW
1p
15 T
440.4 m
103155 kW
CH4 9.896 m
LHV= 495277 kWth
25 T
33 T
74.22 M
90 T
115 p
565 T
59.68 M
87 T
0.05 p
33 T
74.03 M
1.054 p
101 T
27.5 p 565 T
31.52 p 375 T
116.7 p 567 T
4.005 p 264 T
67.15 M
1.497 M
8.566 M
7.058 M 0.0506 M
65.32 M
57.93 M
33 T
FW
LTE
LPE
91 T
450.3 M
LPB
IPE2
LPS
IPB
HPE3
IPS2
HPB1
HPS3
IPE1
1.054 p
90 T
74.22 M
4.286 p 120.8 p 30.45 p 120.8 p
265 T 322 T
320 T 325 T
7.058 M 60.28 M 7.399 M 59.68 M
276
280
335
338
31.67 p 31.67 p
234 T 237 T
146 T
67.15 M 7.402 M
146 T
129
158
196
250
p[bar], T[C], M[kg/s], Steam Properties: Thermoflow - STQUIK
567 T
567 T
480
631
GT PRO10.9 Maher Elmasri
LTE
17717
700
IPE1
12448
LPE
1624.2
LPB
18177
IPE2
26359
Net Power 276262 kW
LHV Heat Rate 6454 kJ/kWh
HRSGTemperature Profile
HPB1
LPS
72240
1785.9
IPS2
IPB
1789.7
12977
HPE3
27046
RH3
29375
HPS3
50008
0
Q kW
600
500
TEMPERATURE [C]
631 T
450.3 M
RH3
4.5 p
148 T
8.555 M
5
400
76
300
8
11
12
14
200
15
17
UA kW/C
100
372.9
0
0
87.93
687.3
890.2
756.6
50
546.9
1112.3
34.27
100
38.05
485.4
339.7
1265.7
150
200
250
300
HEAT TRANSFER FROMGAS [.001 X kW]
Figures 14a (top) and 14b (bottom). Flow schematic and heat recovery temperature profile of the triple-pressure, reheat
combined cycle used in the examples
12-15
 M. A. Elmasri, 1990-2008
Chapter 12: Triple-Pressure & Reheat CC
12.3.4 PRACTICAL CONSIDERATIONS IN REHEAT CYCLE LAYOUT
The reheat cycles shown in Figs. 12 through 14 have the HP superheater laid out in parallel to the reheater.
This allows the smoothest heat recovery temperature profile, corresponding to the smallest HRB surface
area. However, it has some practical drawbacks.
First, if the gas temperature profile entering the boiler is distorted, with one side of the boiler duct receiving
hotter gas than the other, the superheat or reheat temperature desired may not be obtained. Second, placing
the tube bundles in parallel allocates the HRB duct width roughly 60% to reheater tubes and 40% to
superheater tubes. Thus, the reheater tubes, which carry a large volume of steam, would have only 60% as
many tubes per row as they might have had if they were allowed to occupy the entire HRB duct width. This
would result in high steam velocities, or the need to employ a two-row-per-pass or even a three-row-perpass arrangement for the reheater, reducing heat transfer effectiveness. (On the other hand, the HP
superheater, which carries dense, high-pressure steam, benefits from the higher steam velocity of the parallel
layout.)
As a result, most HRB designs for reheat cycles use interleaved superheaters and reheaters as illustrated by
Figures 15a & 15b. This allows each of the sections to occupy the entire duct width. The thermodynamic
performance of the model of Figures 15 and Figures 14 are identical, since the final states of steam leaving
the HRB are identical. However, the required heat transfer surface areas will differ slightly. The
interleaved arrangement requires a larger total (UA).
In designing the profile of Figure 15b, one should note that the temperature rise in each of the interleaved
bundles should be smaller for the hotter bundles than for the cooler ones. This more-closely approximates
the thermodynamically-preferred, parallel arrangement of superheater & reheater.
12-16
 M. A. Elmasri, 1990-2008
Chapter 12: Triple-Pressure & Reheat CC
GT PRO 10.9 Maher Elmasri
Net Power 276263 kW
LHV Heat Rate 6454 kJ/kWh
1.01 p
15 T
60 %RH
440.4 m
0 m elev.
450.3 m
1X GE 7251FB
18.55 p
416 T
17.62 p
1369 T
74.29 %N2
12.16 %O2
3.914 %CO2+SO2
8.742 %H2O
0.8947 %Ar
1.04 p
632 T
450.3 M
181886 kW
1p
15 T
440.4 m
103155 kW
CH4 9.896 m
LHV= 495277 kWth
25 T
33 T
74.22 M
90 T
115 p
565 T
59.68 M
87 T
0.05 p
33 T
74.03 M
1.054 p
101 T
27.5 p 565 T
31.52 p 375 T
116.7 p 567 T
4.005 p 264 T
67.15 M
1.497 M
8.566 M
7.058 M 0.0506 M
65.32 M
57.93 M
33 T
FW
LTE
LPE
91 T
450.3 M
LPB
IPE2
LPS
IPB
HPE3
IPS2
HPB1
HPS0
RH1
HPS1
RH3
HPS3
631 T
450.3 M
IPE1
1.054 p
90 T
74.22 M
4.5 p
148 T
8.555 M
4.286 p 120.8 p 30.45 p 120.8 p 119.4 p 29.46 p 118.1 p 28.46 p 116.7 p
265 T 322 T
320 T 325 T 467 T
507 T 527 T
567 T 567 T
7.058 M 60.28 M 7.399 M 59.68 M 59.68 M 65.32 M 59.68 M 65.32 M 59.68 M
276
280
335
338
480
546
585
603
620
631
31.67 p 31.67 p
234 T 237 T
146 T
67.15 M 7.402 M
146 T
129
158
196
250
p[bar], T[C], M[kg/s], Steam Properties: Thermoflow - STQUIK
GT PRO10.9 Maher Elmasri
LTE
17717
700
IPE1
12448
LPE
1624.3
LPB
18177
IPE2
26359
Net Power 276263 kW
LHV Heat Rate 6454 kJ/kWh
HRSGTemperature Profile
LPS
HPB1
1786
72240
IPB
IPS2
12977
1789.7
HPE3
27046
HPS0
34178
Q kW
600
HPS1
9655
RH1
20532
2
3
HPS3
6174
RH3
8843
0
1
4
TEMPERATURE [C]
500
5
400
76
300
8
11
12
14
200
15
17
UA kW/C
100
372.9
0
0
87.93
687.4
890.2
756.6
50
546.9
1112.3
34.27
100
38.05
303.6
1265.7
150
200
169.2
121.8
79.38
101.1
250
300
HEAT TRANSFER FROMGAS [.001 X kW]
Figures 15a (top) and 15b (bottom). Flow schematic and heat recovery temperature profile of the triple-pressure, reheat
combined cycle used in the examples, but with interleaved, not parallel, superheaters and reheaters
12-17
 M. A. Elmasri, 1990-2008
Chapter 12: Triple-Pressure & Reheat CC
12.3.5 DUAL-PRESSURE REHEAT CYCLE/HIGH PRESSURE REHEAT
One problem of the 2PRH cycle of Fig. 13 is that the IP pressure is constrained to be the same as the reheat
pressure. An optimum reheat pressure is generally much higher than an optimum IP for efficient twopressure heat recovery, both thermodynamically and practically.
Thermodynamically, one would like the IP of a two-pressure system in the range 6-12 bar (90 to 180 psia)
for effective heat recovery. This is too low for reheat from the thermodynamic viewpoint unless reheat
temperature were reduced or condenser pressure very low, otherwise the steam quality at the end of the
expansion line to the condenser would be too high, or even superheated. From the practical viewpoint, low
reheat pressure complicates the reheater physical design. The large mass flow through the reheater at low
pressure results in very large volume flow rate. To accommodate that in typical HRB geometries at
reasonable steam velocities and pressure drops requires either large-diameter tubes (about 4") or multiple
tube rows per flow pass. Large tubes have poor heat transfer due to their smaller surface area/flow cross
sectional area ratio. Multiple tube rows per pass make inefficient use of temperature differences since they
are in crossflow. Mixing the cool steam from the rear tubes with the hot steam from the front ones of the
same pass results in a mixture cooler than from a counterflow exchanger of one row per pass.
One may therefore wish to design a dual-pressure cycle with reheat at a higher pressure than IP. Figure 16
shows an example of such a cycle, designated 2P-HPRH (for High-Pressure Reheat) in Figures 4 through
11. Here, reheat is at 26 bars (377 psia) but IP steam induction is at 7 bars (102 psia). The low IP improves
heat recovery, so the stack is considerably cooler than 2PRH as seen in Fig. 4. The cycle is intermediate
between 2PRH and 3PRH in performance and cost. On performance improvement, 2P-HPRH is closer to
3PRH than it is to 2PRH. To within the uncertainty inherent in estimated cost differences, the cost increase
from 2PRH to 2P-HPRH is less than proportionate to the improved performance. The 2P-HPRH cycle may
therefore be a cost-effective alternative to 3PRH, particularly where the low-temperature stack of 3PRH
needs to be avoided, such as when using oil fuel.
12-18
 M. A. Elmasri, 1990-2008
Chapter 12: Triple-Pressure & Reheat CC
GT PRO 10.9 Maher Elmasri
Net Power 275073 kW
LHV Heat Rate 6482 kJ/kWh
1.01 p
15 T
60 %RH
440.4 m
0 m elev.
450.3 m
1X GE 7251FB
18.55 p
416 T
17.62 p
1369 T
74.29 %N2
12.16 %O2
3.914 %CO2+SO2
8.742 %H2O
0.8947 %Ar
1.04 p
632 T
450.3 M
181886 kW
1p
15 T
440.4 m
101743 kW
CH4 9.896 m
LHV= 495277 kWth
25 T
33 T
73.85 M
100 T
115 p
565 T
61.02 M
87 T
0.05 p
33 T
73.69 M
1.2 p
105 T
0.0724 M
12.82 M
7 p 318 T
26.17 p 565 T
30 p 369 T
116.7 p 567 T
73.97 M
0.6676 M
0.546 M
59.27 M
59.27 M
33 T
FW
LTE
LPB
1.2 p
100 T
73.85 M
1.2 p
105 T
0.6676 M
IPE2
IPB
IPS1
HPE3
IPS2
HPB1
HPS3
95 T
450.3 M
631 T
450.3 M
RH3
7.571 p 7.571 p
165 T 168 T
73.97 M 12.82 M
138
142
181
p[bar], T[C], M[kg/s], Steam Properties: Thermoflow - STQUIK
7.426 p
223 T
12.82 M
235
120.8 p 7.28 p 120.8 p
320 T 325 T
322 T
61.63 M 12.82 M 61.02 M
238
333
338
567 T
566 T
484
631
GT PRO10.9 Maher Elmasri
LTE
20726
700
LPB
1463.1
IPB
26020
IPE2
19028
Net Power 275073 kW
LHV Heat Rate 6482 kJ/kWh
HRSGTemperature Profile
IPS1
HPB1
1650.5
73865
IPS2
2650.2
HPE3
46642
RH3
26613
HPS3
51132
500
TEMPERATURE [C]
0
Q kW
600
5
400
76
300
9
11
200
12
14
17
UA kW/C
100
422.6
0
0
769.6
41.81
796.3
50
1490.1
52.35
47.87
100
490.7
302.5
1276.4
150
200
250
300
HEAT TRANSFER FROMGAS [.001 X kW]
Figures 16a (top) and 16b (bottom). Flow schematic and heat recovery temperature profile of the dual-pressure, reheat
combined cycle with reheat pressure above the IP
12-19
 M. A. Elmasri, 1990-2008
Chapter 12: Triple-Pressure & Reheat CC
12.3.6 DUAL-PRESSURE "POOR MAN'S REHEAT" CYCLE
This is illustrated in Figure 17 and is not truly a reheat cycle, rather a weak "reheat effect" is accomplished
by mixing very hot IP steam into the steam turbine expansion path, at the cross-over between the steam
turbine’s HP/IP and LP casings. The cross-over pipe would have to be equipped with a mixing system to
ensure thorough blending of the hot steam as it is added to the cooler steam exiting the HP/IP casing.
Typically, the HP/IP steam would be warmed up by about 35-40 °C (63-72 °F) after blending the hot IP.
This has been dubbed “poor man’s reheat”, and its results are shown in Figures 4 through 11 with the
designation 2P-PMRH.
The performance advantage is quite modest when compared with 2P at the same HP throttle and condenser
pressures, as seen in Figs. 4 through 11.
However, this weak "reheat effect" enables the HP throttle pressure to be raised, and/or the condenser
pressure to be reduced, without running into excessive steam turbine exhaust moisture. Moisture reduces
steam turbine efficiency. Additionally, if the moisture is too high, above 10% say, it can cause blade
erosion unless it is separated and drained off, as is common in nuclear cycle steam turbines.
“Poor Man’s Reheat” is an elegant solution that allows a larger, multi-casing steam turbine to provide better
efficiency by employing a high throttle pressure and a low condenser pressure, if the ambient and economics
permit it, without true reheat, and without the complexity of a steam turbine moisture removal system. As
such it is worthy of consideration in special circumstances.
12-20
 M. A. Elmasri, 1990-2008
Chapter 12: Triple-Pressure & Reheat CC
GT PRO 10.9 Maher Elmasri
Net Power 270742 kW
LHV Heat Rate 6586 kJ/kWh
1.01 p
15 T
60 %RH
440.4 m
0 m elev.
450.3 m
1X GE 7251FB
18.55 p
416 T
17.62 p
1369 T
74.29 %N2
12.16 %O2
3.914 %CO2+SO2
8.742 %H2O
0.8947 %Ar
1.04 p
632 T
450.3 M
181886 kW
1p
15 T
440.4 m
97205 kW
CH4 9.896 m
LHV= 495277 kWth
25 T
33 T
79.96 M
95 T
80 p
565 T
71.56 M
87 T
0.05 p
33 T
79.93 M
1.2 p
105 T
0.0141 M
81.6 p 567 T
8 p 549 T
80.03 M
1.473 M
1.404 M
8.389 M
33 T
FW
LTE
LPB
IPE2
1.2 p
95 T
79.96 M
1.2 p
105 T
1.473 M
8.653 p
171 T
80.03 M
IPB
IPS1
HPE2
HPE3
HPB1
HPS0
IPS2
HPS1
HPS3
631 T
450.3 M
89 T
450.3 M
8.653 p 85.72 p 8.486 p 84.46 p
174 T 280 T
284 T 296 T
8.389 M 72.28 M 8.389 M 72.28 M
187
221
295
299
133
140
p[bar], T[C], M[kg/s], Steam Properties: Thermoflow - STQUIK
84.46 p 83.5 p 82.55 p 8.32 p
550 T
299 T 310 T 450 T
71.56 M 71.56 M 71.56 M 8.389 M
312
513
522
584
GT PRO10.9 Maher Elmasri
LTE
20752
700
HRSGTemperature Profile
HPE3
6182
IPS1
2117.7
HPE2
36122
IPB
16740
IPE2
22508
LPB
3281
Net Power 270742 kW
LHV Heat Rate 6586 kJ/kWh
HPS1
32359
HPS0
4636
HPB1
102415
Q kW
600
81.6 p
567 T
71.56 M
593
631
HPS3
20600
IPS2
4743
0
2
3
4
5
TEMPERATURE [C]
500
400
98
300
7
10
12
200
1714
UA kW/C
100
447.8
0
0
104.7
935.7
1275.5
631.3
50
41.85
353.1
100
1431.8
21.84
191.3
150
200
36.14
209.8
250
300
HEAT TRANSFER FROMGAS [.001 X kW]
Figures 17a (top) and 17b (bottom). Flow schematic and heat recovery temperature profile of the dual-pressure, “poorman’s reheat” combined cycle
12-21
 M. A. Elmasri, 1990-2008
Chapter 12: Triple-Pressure & Reheat CC
12.4
EFFECT OF SUPPLEMENTARY FIRING ON REHEAT COMBINED CYCLES
As discussed in §12.3.1, the 1PRH cycle can have a duct burner fuel utilisation efficiency around 155%, and
its internal efficiency is very high, 39.5% in our example. Thus, the efficiency of converting duct burner
fuel to electricity can be ≈ 1.55*.395 ≈ 61%, as long as the incremental supplementary firing can be used to
depress stack temperature. This is higher then the base CC efficiency, which is in the mid 50’s. Hence,
supplementary firing can improve the efficiency of a single-pressure reheat cycle, as long as the design is in
the range where incremental duct firing can lower the stack temperature. Once the stack temperature is at its
minimum acceptable value, however, further supplementary firing can only be converted to electricity at the
bottoming cycle internal efficiency of 39.5%, and will thus degrade the net plant efficiency.
The dual-pressure and triple-pressure cycles have lower internal efficiencies. They also have lower duct
burner fuel utilisation efficiencies, since the rate of stack temperature depression per unit of inlet gas
temperature rise is lower. Hence their efficiency does not benefit from supplementary firing, as seen in
Figure 18. It should be noted, however, that supplementary firing improves the bottoming cycle internal
efficiency of the multi-pressure cycles, since it results in a greater percentage of the heat going to the HP
than to the IP or LP. Indeed, as supplementary firing level is raised, the internal efficiency of a multipressure cycles increases, until it equals that of the single-pressure cycle when its lower pressures have been
eliminated from the design. The discontinuities in slope in Fig. 18 at 675 °C duct firing for the 3PRH, and
at 700 °C duct firing for the 2PRH and 1PRH, are due to the steam turbine design changing from a single to
a double flow LPT (with lower exhaust loss), to accommodate the increasing exhaust volumetric flow rate.
Effect of supplementary firing on reheat combined cycle efficiency
56.00
Net combined cycle efficiency, %
3PRH
2PRH
55.50
1PRH
LP eliminated
55.00
IP eliminated, Tstack, min of
90 °C reached for 1PRH
54.50
54.00
53.50
625
650
675
700
725
750
775
800
Duct burner exit temperature, C
Figure 18. Effect of supplementary firing on the reheat combined cycles used for illustration
Supplementary firing beyond the point that attains minimum stack temperature is most effective if feedwater
heaters are introduced in parallel with the economiser of the single-pressure combined cycle. However,
even with this measure, net combined cycle efficiency still falls with further supplementary firing, since the
maximum internal efficiency of a conventional steam cycle with feedwater heaters is around 46%, for a
seven-heater subcritical cycle, and around 50%, for a double-reheat supercritical cycle. This represents the
upper bound on converting duct burner fuel to electricity once minimum stack temperature has been
attained, and, since it is lower than the combined cycle efficiency, it will drag it down. The most efficient
way to convert fuel to electricity is to burn it in the gas turbine, since the working fluid receiving this energy
is at a much higher temperature than the steam in a bottoming cycle, as explained in Chapter 1.
12-22
 M. A. Elmasri, 1990-2008
Chapter 13: Performance Enhancement
PERFORMANCE ENHANCEMENT
Revised September, 2008
© Maher Elmasri 1990-2008
13.1
INTRODUCTION & GENERAL OBSERVATIONS
In this section we shall describe, with quantitative examples, the various design features that can be adopted
to enhance performance of a gas turbine based power plant. These enhancements usually require installation
of additional equipment, which adds cost and, in most cases, parasitic power losses. The extra equipment
may be thermodynamically coupled to the gas turbine alone, the steam cycle alone, or integrated with both.
Several of these enhancements act primarily on the gas turbine, to increase its power output through
increased airflow. These enhancements can increase both the power output and the efficiency of the gas
turbine on its own. If the gas turbine is part of a combined cycle plant, the general rule is that combined
cycle output will also increase, but less than proportional to the increase in gas turbine output, and combined
cycle net efficiency will be very little affected and may even decrease. This is a manifestation of the old
adage, “no free lunch”. Any improvement in gas turbine efficiency means that its exhaust energy will be
reduced as a percentage of its fuel input. If the bottoming cycle converted exhaust energy to electricity at a
constant efficiency, then overall plant efficiency need not decline, but the bottoming cycle may not maintain
its efficiency. The quality of the gas turbine exhaust energy declines, because to produce more power and
become more efficient, it must have a higher airflow and a lower exhaust temperature, if its fundamental
cycle is unchanged 1 and its firing temperature is not increased. The combination of higher exhaust flow and
lower exhaust temperature means that the exergy/energy ratio of the exhaust is reduced, reducing the
percentage of its energy which can be converted to power with fixed equipment. If the bottoming cycle
equipment could be re-designed (enlarged) to handle the higher mass flow of cooler gases, and to still
convert their energy to power at the same efficiency, than overall plant efficiency may increase slightly, but
so will capital cost. If the bottoming cycle equipment were physically the same, with and without the gas
turbine enhancement, then the higher exhaust flow and lower temperature will degrade its efficiency, and the
combined cycle net efficiency will decline. In addition to these fundamental facts, the extra equipment to
enhance gas turbine performance introduces additional parasitic losses, either as pressure drops or as
auxiliary power consumption, or both. Examples of these sorts of enhancements include inlet air
evaporative cooling and fogging (§13.2), refrigeration (§13.3), and supercharging (§13.4).
Design enhancements which improve overall combined cycle net efficiency will, generally, reduce its power
output. This arises from the fact that most efficiency enhancements are regenerative, using low-grade heat
to save fuel rather than to generate additional power. Examples are fuel heating in combined cycles using
low-grade hot water from the boiler, discussed in §13.5, or regenerative feedwater heating in conventional
steam plants. These measures improve efficiency by using low-grade energy to conserve fuel (high grade
energy), but reduce output per unit of primary working fluid, such as from a gas turbine of given airflow or
from a steam turbine of given throttle flow.
Overspray fogging is the only enhancement discussed in this chapter that introduces a fundamental change to the gas
turbine cycle
1
13-1
 M. A. Elmasri, 1990-2008
Chapter 13: Performance Enhancement
13.2
INLET-AIR EVAPORATIVE COOLING & FOGGING
13.2.1 EVAPORATIVE COOLING (OR UNDERSPRAY FOGGING)
13.2.1.1 Wet-Media Evaporative Cooling
Evaporative cooling has traditionally been accomplished by having the air flow in intimate contact with wet
media. Vertical porous pads inserted into the inlet structure are kept moist by a continuous water flow
trickling down through them. The air goes through drift eliminators, immediately following the wet media,
to coalesce and drain the entrained, unevaporated liquid droplets. Surplus water is collected in a basin at
their bottom and recirculated, with makeup, via a pump to their top.
Due to the intimate contact between air and water, the air is adiabatically saturated, but not quite to 100%
relative humidity. Since attaining complete saturation would require the media to be either of very tight
mesh or very thick, the air is not totally saturated. The effectiveness, ε, of an evaporative inlet cooler is
commonly defined as the temperature drop experienced by the air relative to the temperature difference
between the ambient dry-bulb and wet-bulb temperatures:
ε=
Tamb − Tc
Tamb − Twb
.......................... (1)
where Tamb is the ambient dry-bulb temperature, Tc is the temperature of the cool air leaving the evaporative
cooler, and Twb is the ambient wet-bulb temperature. Practical installations have an effectiveness of between
85% and 95%.
13.2.1.2 Fogging
More recently, an alternative procedure, known as fogging, is gaining popularity. In this method, the water
is atomised by forcing it at high pressure (about 2500 psi or 170 bar) through minute holes (about 5 μm in
diameter) in spray nozzles. An array of these spray nozzles is fitted into the inlet plenum of the gas turbine
to distribute the atomised water and mix it thoroughly with the air, where it evaporates. If the flow rate of
water is controlled such that it all evaporates, cooling the air to essentially its wet-bulb temperature, the
thermodynamics are very similar to the media type evaporative cooler. Fogging has the thermodynamic
advantage that the air pressure drop is essentially zero, whereas a media-type evaporative cooler and its
drift-eliminator can add 1 mb (0.4 “H2O) to the inlet pressure drop. Figure 1 shows this type of system.
Wet-media evaporative cooling can utilise ordinary city water, by relying on blowdown to limit the
concentration of salts in the system, but demineralised water is preferable, to avoid or reduce blowdown and
media replacement. Fogging requires demineralised water, since any salts emerge as a fine dust as the fog
evaporates, and deposit on the air ducts and the gas turbine inlet.
Fogging equipment is somewhat more complex than wet-media systems. Wet-media coolers will naturally
evaporate the right amount of water, as long as they are kept wet, and surplus water will trickle down the
media and be recycled. A drift eliminator captures most liquid carryover and recycles it. Fogging systems
must inject the right amount of water based on measured inlet conditions and measured or calculated engine
airflow. Excess water injection will carry over into the engine, since the water particles are very fine. If that
is acceptable, a little overspray will only enhance performance, but some manufacturers’ warranties may be
violated if routine overspray of fog into the engine is practiced. Once the right flow rate of fog is calculated
at any operating condition, its implementation is complicated by the fact that the spray nozzles have a very
limited range of turndown, since any reduction in water pressure to reduce flow rate results in coarser
atomisation. Hence, varying the water injection flow rate requires turning on or off entire manifolds of
nozzles, and the system needs to be designed with many manifolds, each supplying a different number of
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Chapter 13: Performance Enhancement
nozzles, so that various permutations and combinations of the manifolds in use can cover a wide range of the
active number of nozzles and thus of water flow rates.
Figure 1. Fogging nozzles and manifolds in the gas turbine’s inlet air duct (left), and spray from a nozzle (right).
Courtesy of Mee Fogging Systems.
13.2.1.3 Thermodynamic Effects of Evaporative Inlet Cooling
The latent heat of water at 25 °C (77 °F) is about 2440 kJ/kg (1050 BTU/lb). The specific heat of air is
about 1 kJ/kg-°C (0.24 BTU/lb-°F). Thus, if 0.4% by mass of water were mixed and evaporated into air, the
air would cool down by about 10 °C (18 °F). Since a typical gas turbine’s output increases by ≈7% per 10
°C of inlet air temperature reduction (§6.3.1.4), the rate of increase of a typical gas turbine’s output is on the
order of 17% output gain per 1% water evaporation, by mass relative to airflow. By contrast, when steam
injection is employed to increase power output, a typical gas turbine gains about 3% in power output per 1%
of steam injection, by mass relative to airflow. Thus, as far as demineralised water consumption is
concerned, evaporative cooling is more effective than steam injection, by a factor of five, for a given gas
turbine power gain. Compared with steam injection, evaporative cooling has the additional benefit that
when T1 is reduced, T2 is also reduced, which helps extend hot section life because the turbine blades
receive cooling air at a lower temperature.
The main drawback of evaporative cooling is that it is ineffective when the ambient is too humid or too cool,
since the amount of water that can be evaporated depends on the difference between the dry-bulb and wetbulb temperatures. This difference is significant in hot/dry climates, but small in cool/humid climates.
Nevertheless, under many circumstances, evaporative cooling is one of the most cost-effective means of
increasing power output. In a warm ambient, around 30 °C (86 °F), with medium relative humidity, around
50%, evaporative inlet cooling or underspray fogging can add to a combined cycle’s capacity at a capital
cost of about $200-300/incremental kW, depending on plant size, including the additional costs of water
treatment, tanks, piping, engineering, field labour, etc. Since this is much less than the cost of the base
combined cycle, it implies a reduction in the total specific cost ($/kW), typically by 2-3%.
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Chapter 13: Performance Enhancement
Example of the Effect of Evaporative Inlet Air Cooling on GT Performance
50
Effect of a 90%-effective Evap Cooler, %
9
59
68
77 F
86
95
104
30
35
40
GT Power, 40% RH
8
GT Power, 80% RH
7
GT Heat Rate, 40% RH
GT Heat rate, 80% RH
6
5
4
3
2
1
0
-1
-2
-3
10
15
20
25
Ambient Temperature, C
Figure 2. Effect of evaporative inlet cooling on a typical 40-MW-class gas turbine’s performance as a function of
ambient, for two levels of ambient humidity: 40% and 80%.
Figure 2 illustrates the effect of an evaporative cooler on the performance of a typical, heavy duty gas
turbine. The curves show the power gain and heat rate improvement as a function of ambient temperature,
at two levels of ambient relative humidity, 80% and 40%. The benefit of the evaporative cooler is greatest
when the ambient is warm and dry. At any ambient temperature, the power gain by evaporative cooling in a
40%-humid ambient is about triple that in an 80%-humid ambient.
Example of the Effect of Evaporative Inlet Air Cooling on CC Performance
50
Effect of a 90%-effective Evap Cooler, %
7
59
68
77 F
86
95
104
30
35
40
CC Power, 40% RH
6
CC Power, 80% RH
5
CC Heat rate, 80% RH
CC Heat Rate, 40% RH
4
3
2
1
0
-1
10
15
20
25
Ambient Temperature, C
Figure 3. Effect of adding evaporative inlet cooling to a typical 60-MW-class dual-pressure combined cycle, with fixed
combined cycle plant equipment.
Figure 3 shows the effects of evaporative inlet cooling on a typical dual-pressure combined cycle. The
power gain for the whole plant is lower on a percentage basis than for its gas turbine. Although the gas
turbine’s heat rate improves, as seen in Figure 2, the heat rate for the combined cycle is essentially
unaffected, and indeed worsens slightly in our example.
To explain why evaporative cooling is less beneficial to the combined cycle than to the gas turbine, we first
note that the principal impact of cooling the gas turbine inlet air is a larger mass flow rate of exhaust gases at
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Chapter 13: Performance Enhancement
a lower exhaust temperature. As explained in Chapter 11, both of these effects are detrimental to the heat
recovery boiler’s efficiency, resulting in a higher stack temperature, a lower percentage of the exhaust
energy being recovered, and a lower steam temperature at the steam turbine’s throttle. The latter reduces the
internal efficiency of the steam bottoming cycle, exacerbating the reduction in heat recovery efficiency.
These effects are shown in Figure 4 for the same numerical example of Figures 2 and 3, except that for
clarity, only the results for the 40%-humid ambient are shown.
Example of the Effect of Evaporative Inlet Air Cooling on Bottoming Cycle
50
59
68
77 F
86
95
104
25
30
35
40
7
Effect of a 90%-effective Evap Cooler, %
Fuel Input
Exhaust Energy
6
Recovered Energy
ST Output
5
ST Pseudo HR
4
3
2
1
0
10
15
20
Ambient Temperature, C (40% RH)
Figure 4. Effects of evaporative inlet cooling on a typical dual-pressure bottoming cycle (fixed equipment)
The top curve in Figure 4 shows the change in plant fuel input occasioned by the evaporative inlet cooler.
This is the same as the gas turbine fuel input in the absence of a duct burner, which is the case in this
example. Because the gas turbine’s efficiency improves with cooler air, its exhaust energy does not increase
in proportion to its fuel input, as shown. Furthermore, because the heat recovery boiler is less effective
when supplied by a higher mass flow rate of cooler exhaust gas, the recovered heat does not increase in
proportion to the exhaust energy. Finally, because the steam turbine throttle temperature falls, its power
output does not increase in proportion to the recovered heat.
The increase in steam turbine output with evaporative cooling is quite meager in relation to the increase in
gas turbine output or gas turbine fuel input. Although the proportion of the plant fuel input that emerges as
gas turbine power increases, the proportion of the plant fuel input that emerges as steam turbine power
declines. This is shown by the curve labelled “ST Pseudo HR” in Figure 4. We define the steam turbine
pseudo heat rate as
ST _ Pseudo _ HR =
Plant _ Fuel _ Input
ST _ Output
.............. (2)
which is a very useful concept in understanding combined cycle behaviour. In our example, Figure 4 shows
that at 40 °C and 40% humidity, the evaporative inlet cooler makes the steam turbine pseudo heat rate worse
by just over 4%. Figure 2 shows that under these conditions, the evaporative inlet cooler it makes the gas
turbine heat rate better by about 2%. Since the steam turbine output is about half the gas turbine’s, the net
result on overall plant heat rate is nearly neutral, as shown in Figure 3. In effect, the decline in steam
turbine pseudo heat rate with evaporative cooling cancels the improvement in gas turbine heat rate.
The results shown in Figures 2 through 4 are off-design simulations, obtained with the premise that the
power plant equipment is fixed, having been designed at some nominal set of conditions, and is being
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Chapter 13: Performance Enhancement
operated at different ambients and humidities with the evaporative cooler either on or off. When the cooler
is on, the gas turbine produces a larger mass flow of cooler exhaust gases, and the fixed bottoming cycle
components perform less efficiently. Figure 5 is the equivalent of Figure 3, but on the premise that the
bottoming cycle equipment is redesigned to obtain the same steam cycle main design parameters, such as
throttle pressure and temperature, at each set of exhaust gas conditions. In this design calculation, a higher
mass flow of cooler exhaust implies a boiler redesigned with larger surface areas to obtain the same pinch
points, the same steam temperatures, and the same pressure drops, i.e. a more expensive boiler. It implies a
larger steam turbine and a larger condenser, designed to accommodate the larger steam flows arising from
the larger exhaust gas flow, occasioned by evaporative cooling. As would be expected, the design
calculations of Figure 5 shows a greater thermodynamic benefit than the fixed-equipment calculations of
Figure 3. In particular, they show that combined cycle heat rate can actually improve, slightly, with
evaporative inlet cooling.
Example of the Effect of Evaporative Inlet Air Cooling on CC Performance if
Bottoming Cycle Equipment were Resized for the Changed Exhaust Conditions
Effect of a 90%-effective Evap Cooler, %
7
50
59
68
77 F
86
95
104
25
30
35
40
CC Power, 40% RH
6
CC Power, 80% RH
5
CC Heat rate, 80% RH
CC Heat Rate, 40% RH
4
3
2
1
0
-1
10
15
20
Ambient Temperature, C
Figure 5. Effect of evaporative inlet cooling on a typical 60-MW-class dual-pressure combined cycle, if the bottoming
cycle components were resized to accommodate the altered exhaust conditions resulting from evaporative cooling
The results shown in Figure 5 evaluate the impact of evaporative cooling with larger, more expensive
bottoming cycle components than would have been purchased in its absence. In these results, the cost of
adding evaporative cooling has two components, the cost of the evaporative cooling equipment itself, and
the additional costs associated with larger bottoming cycle equipment. The results shown in Figure 3
evaluate the impact of evaporative cooling with the same bottoming cycle equipment in place. In these
results the only additional costs are those associated with the evaporative cooling system.
Whenever the benefits of such a performance enhancing measure are discussed, it is important to stipulate
the premise of the calculations, whether the enhancement is appended to fixed equipment, such as in the
results of Figures 2-4, or whether the enhancement involves re-sizing other equipment to maximise its
potential benefits, such as in the results of Figure 5. Both these types of calculations are perfectly legitimate
at the design phase of a yet unbuilt power plant, but only the first type is meaningful for a retrofit to an
existing power plant.
13.2.2 OVERSPRAY FOGGING
A fogging system can be characterised as either underspray or overspray. An underspray system is one
where the atomised water flow is controlled to be just equal to the amount that can be fully evaporated by
the inlet airflow, based on its ambient temperature and humidity. Thermodynamically, this is virtually
identical to the wet-media type of evaporative cooling. An overspray system is one where the amount of
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Chapter 13: Performance Enhancement
atomised water is set to be greater than that which may be evaporated by the inlet airflow, resulting in some
liquid fog carrying into the gas turbine compressor. There is some controversy about overspray fogging,
due to concerns on possible erosion of the compressor blades due to impact with water particles, although
these particles are designed to be very small, on the order of 10-20 μm. Small particles, below about 10 μm,
tend to follow the air streamlines, rather than slip out and strike the blades.
500
500
400
400
300
300
W=409 kJ/kg
200
100
0
Temperature [°C]
Temperature [°C]
16.21[bar]
4
5
6
7
Specific Entropy
8
W=385 kJ/kg air
(6% reduction)
200
100
1.013[bar]
3
16.21[bar]
9
0
1.588[bar]
1.013[bar]
2
3
4
5
Specific Entropy
6
7
Figure 6. T-s diagrams for dry (left) and wet (right) compression, beginning at ISO conditions, at a polytropic efficiency
of 90%, and with a pressure ratio of 16
From the thermodynamic perspective, however, overspray fogging is very effective. Figure 6 shows T-s
diagrams of a normal, dry compression and a wet compression. Both processes begin at 1.013 bar/15 °C
(14.696 psia/59 °F) with 100%-humid air, and both have a pressure ratio of 16. In the wet compression, the
air carries entrained liquid fog into the compressor inlet, at a rate of 1% of air mass flow rate. As the air is
adiabatically compressed, it heats up and the water evaporates, cooling the air. After all the water has
evaporated, at about 1.6 bar in the example, the rest of the compression is all in the gaseous phase. With 1%
overspray, the fog all evaporates by the third or fourth compressor stage. The intercooling effect results in a
considerable reduction in compressor work, about 6% in this example. Additionally, the turbine generates
about 1.5% more work, since it receives 1% more gases by mass, and about 1.5% more gases by volume or
moles, which is what matters in calculating work output. Since, in a typical gas turbine, the compressor
absorbs about half the turbine’s work, decreasing the compressor work by 6% and increasing the turbine
work by 1.5% result in an increase in net engine output of about 9% {(101.5-47)/(100-50) = 1.09}.
With wet compression, the compressor discharge temperature, T2, drops. In our example, 1% overspray
reduces T2 by about 50 °C (90 °F). This necessitates additional fuel input to the combustor to attain the
same value of T3 as without overspray. Since, in a typical gas turbine, the combustor temperature rise (T3T2) is on the order of 750-900 °C, the additional fuel input is on the order of 6%. Therefore 1% overspray
results in ≈9% more power output and requires ≈6% more fuel input, and so improves heat rate by ≈3%.
Naturally, the precise results differ from one gas turbine model to another, with the benefit being stronger
for machines with a poor net work ratio (net work to turbine work), in which the compressor absorbs a
greater proportion of the turbine’s output. Low net work ratio machines are those with a relatively high
pressure ratio and/or a relatively low firing temperature.
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Chapter 13: Performance Enhancement
Another, but secondary, effect of wet compression is an increase in pressure ratio, and attendant decrease in
exhaust temperature. An overspray of 1% (by mass) increases the pressure ratio by about 1.5% (the increase
in mole flow rate entering the turbine). This, on its own, would cause an exhaust temperature drop on the
order of 0.3% (assuming the turbine expansion exponent xT ≈ 0.2), or about 2.5 °C (4.5 °F). However, the
increased water vapour content tends to decrease the value of the turbine expansion exponent, xT. This
partially negates the drop in exhaust temperature if the firing temperature were held constant by a control
system that compensates for it. If the control system does not compensate for increased water vapour
content, then exhaust temperature will be drop as suggested by the higher pressure ratio, and firing
temperature will be inadvertently reduced.
If the gas turbine were part of a combined cycle, wet compression will increase its exhaust gas heat capacity
flux, the product of mass flow rate and specific heat. This, along with the lower exhaust temperature, both
reduce the efficiency of a given heat recovery boiler, and hence of a steam bottoming cycle of fixed design.
Figure 7 illustrates the results for a typical, 40-MW-class heavy duty gas turbine, coupled to a dual-pressure
combined cycle. These results are with the premise of fixed bottoming cycle equipment. 100% flow on the
x-axis represents a total fog input of 1.25% of nominal engine air mass flow rate. To the left of the diagram,
up to the 15% fogging flow point, the system is in underspray mode, with the water evaporating before
entering the gas turbine. Beyond that point, where the curves change slope, the system is in overspray
mode, with water carrying into the compressor. The diagram shows that with an overspray of only ½ % of
airflow, the gas turbine output increases by about 11%. Figure 8 shows that the corresponding increase in
gas turbine fuel consumption is 8%, and the corresponding gas turbine heat rate improvement is thus 3%.
Typical 40-MW-class gas turbine with dual-pressure combined cycle, wet cooling tower, ambient 95 F, 70% RH
100% Fogger flow = 1.25 % of GT nominal air mass flow rate
16
14
Increase in output, %
12
underspray
overspray
10
Plant net output
8
GT gross output
ST gross output
6
4
0.5 % overspray
2
0
0
10
20
30
40
50
60
70
80
90
100
Fogger water flow, % of nominal
Figure 7. Effect of fogging on gas turbine, steam turbine, and combined cycle power output
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Chapter 13: Performance Enhancement
Typical 40-MW-class gas turbine, dual-pressure combined cycle, wet cooling tower, ambient 95 F, 70% RH
100% Fogger flow = 1.25 % of nominal air mass flow rate
Change in Heat Rate & GT fuel, %
10
underspray
overspray
8
6
Plant net HR
4
GT gross HR
ST pseudo HR
GT fuel input
2
0
0.5 % overspray
-2
-4
0
10
20
30
40
50
60
70
80
90
100
Fogger water flow, % of nominal
Figure 8. Effect of fogging on gas turbine heat rate, steam turbine pseudo heat rate, and combined cycle heat rate
Figures 8 and 9 illustrate that overspray fogging in the gas turbine also increase steam turbine output, but
very weakly. The cooler gas turbine inlet air increases airflow and exhaust flow, but also causes a reduction
in exhaust temperature. Both effects tend to raise the final stack temperature leaving a HRSG of fixed
surface area, as discussed in Chapter 12 and shown in Figure 9 for this example. Thus, the increase in
HRSG steam production and ST output is very modest. The ST pseudo heat rate, described in Equation (2),
increases, with the result that the net heat rate of the combined cycle is barely affected, as shown in Fig. 9.
Typical 40-MW-class gas turbine, dual-pressure combined cycle, wet cooling tower, ambient 95 F, 70% RH
100% Fogger flow = 1.25 % of nominal air mass flow rate
4
Change in gas temperatures, F
2
0
-2
GT Exhaust Temp.
-4
HRSG Stack Temp
-6
0.5 % overspray
-8
underspray
overspray
-10
-12
0
10
20
30
40
50
60
70
80
90
100
Fogger water flow, % of nominal
Figure 9. Effect of fogging on gas turbine exhaust and combined cycle stack temperatures
One of the collateral benefits of overspray fogging is the reduction of compressor discharge temperature, T2,
which provides the turbine blades with cooler cooling air. This can significantly extend hot-section life,
especially in hot climates, where without fogging their life would be adversely affected by the hotter cooling
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Chapter 13: Performance Enhancement
air. Overspray fogging is a very cost-effective way of improving power output, year round, except under
freezing conditions. Depending upon plant size and site parameters, it can add significant capacity to a
combined cycle at a cost on the order of $300-$400/incremental kW, including additional costs of water
treatment, tanks, piping, engineering, field labour, etc. This a little more expensive than capacity gained by
evaporative cooling in warm, dry ambients, but then, evaporative cooling is only effective under these
conditions whereas overspray fogging is effective under all conditions above freezing temperatures. The
biggest advantage to overspray fogging over evaporative cooling is that it is effective even if the ambient is
hot and nearly 100%-humid, when the power boost is most needed to cope with high air-conditioning loads.
Despite its benefits, overspray fogging is still a subject of controversy, with some gas turbine manufacturers
opposing its use due to concerns about compressor erosion, as well as due to its interaction with DLN
combustor systems, which are sensitive to variations in compressor discharge temperature. Additional
concerns arise from the interaction between overspray fogging and the engine’s controls. For instance if
overspray fogging were turned off suddenly, there would be a sudden rise in T2. The control valve would
still be supplying fuel at a rate that would heat the air from the much lower T2, extant before the fogging
shutoff, to the rated T3. This fuel flow, with the sudden jump in T2, would cause a corresponding sudden
jump in T3, which would last a fraction of a second, until the hotter exhaust propagated through the machine
to the exhaust thermocouples, which would then signal the controls to reduce fuel flow (or trip the machine).
Thus, the controls on a fogging system should only allow the fogging rate to increase or decrease gradually,
to allow the engine’s controls time to accommodate the changes without any thermal shocks.
13.2.2.1 Interstage Fogging
Figure 10. Model of a gas turbine with fog injection at an intermediate location along the compressor
One manufacturer of aeroderivatives is offering machines modified with a fogging spray after a few
compressor stages (the GE LM6000 Sprint). This has the advantage that the fog is injected into hot air, so
that it evaporates quickly, lessening the risk of erosion. It provides the thermodynamic advantages listed
above, but their effect is not quite as strong as fogging right at the inlet, since the greater the proportion of
the compression that benefits from the cooling effect, the better.
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Chapter 13: Performance Enhancement
Effect of fogging location on a GT cycle with a PR of 16
Power & Heat rate, relative to fogging at inlet
103
102
101
Heat rate, relative to
fogging at inlet
100
Power output, relative to
fogging at inlet
99
98
97
96
1
2
3
4
5
6
Fogging pressure/inlet pressure
Figure 11. Performance of the gas turbine model of Figure 10 as a function of fog injection location
To numerically illustrate this principle, a model of a gas turbine with a pressure ratio of 16 has been
constructed, as shown in Figure 10. Fog with a mass flow rate of 0.5% of inlet airflow is injected. The
results of varying the injection location, defined by pressure as a multiple of inlet (ambient) pressure, is
shown in Figure 11. The deeper the injection point into the compressor, the poorer the power output and
heat rate. Similar results are obtained for a pressure ratio of 30, similar to that of the LM6000 Sprint.
13.3
GT INLET AIR REFRIGERATION
Chilling the inlet air of a gas turbine is accomplished by a refrigeration system of the type used for airconditioning of large buildings. Those units are mass-produced and reliable. They produce chilled water
that can be circulated to air-conditioning heat exchangers throughout a building complex, or through a coil
of finned tubes placed in the inlet housing of a gas turbine. There are basically two types of chillers,
electrically-driven and heat-driven. The former are normally called "centrifugal", since large refrigeration
systems usually employ a centrifugal compressor in a vapor-compression refrigeration cycle. The latter are
usually called "absorption" since they operate on the absorption cycle which utilizes heat of solution of a
refrigerant, such as Lithium Bromide.
Chillers used in power plants are usually packaged in a more rugged fashion than those for building use.
The plant environment usually requires enclosed ‘explosion-proof’ motors, weather resistant packaging, and
premium quality piping and other components. Thus, a plant package is generally more expensive than one
for building air conditioning.
Another difference between chillers for gas turbine inlet cooling and those for building air conditioning
arises from the different thermodynamic requirements. In a building application, the chilled water range is
usually quite narrow, about 5-7 °C (9-13 °F), with chilled water supplied to the building at about 5 °C (40
°F) and returning to the chiller at about 11 °C (52 °F). In a power plant, such a narrow range would require
an excessive flow rate of chilled water circulation, with massive pipes and pumps, and the optimum range
can be more than double that for building applications. This requires a different vapour-compression cycle,
with a higher pressure ratio than typical building units. Another solution is to stack chiller units in series,
sequentially chilling the water, doubling or tripling the range. This also has thermodynamic benefits, since
the chiller which first handles the warm return water will have a higher COP, since it will operate with a
lower temperature difference between the load and its heat rejection to the cooling water.
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Chapter 13: Performance Enhancement
The finned-tube coil built into the air intake system is usually made of brass tubes with aluminum fins. The
tubes are arranged into modular, horizontal bundles, to allow condensate to drip into a gutter at the bottom
of each bundle. The tubes are usually about 25 mm (1”) in diameter, and the fin density is quite high,
around 450/m (11/inch). Chilled water/ethylene glycol solution flows within the tubes of the tubes, and the
air passes outside them. Drift eliminators may be installed downstream of the coil to coalesce and remove
any condensed water droplets entrained in the air. Depending on temperature difference, the coils can be
fairly deep, with as many as fifteen rows of tubes. The air-side pressure drop is usually on the order of 25
mb (1” H2O).
Figure 12. Combined cycle with inlet air refrigeration, centrifugal chiller packages with their cooling towers atop each
are in the right foreground. Note size of chilled water piping (courtesy of The Stellar Group)
Centrifugal chillers for gas turbine applications would usually have an installed COP (Coefficient of
Performance) of about 4.5. This means that they will provide 4.5 units of cooling for each unit of electric
energy consumed and would need to reject 5.5 units of energy to a cooling system. Thus, their heat
rejection is about 55/45=1.22 times refrigeration capacity. Their cost is about $300/TOR 2 for the chiller
skid, and about $1000/TOR for the entire system installed, with gas turbine inlet cooling coil, auxiliary
cooling tower, piping, pumps, electrical, controls, foundations, engineering and site labour.
Absorption chillers for gas turbine applications would usually be of the two-stage type, which would require
steam at about 135 psia (9 bar) to drive them. They would have a COP of about 1.1, providing 1.1 units of
cooling for each unit of steam energy consumed. They would need to reject 2.1 units of energy per unit of
steam used. Thus heat rejection to their cooling system is about 2.1/1.1= 1.9 times refrigeration capacity.
Their cost is about $500/TOR for the chiller skid, and about $1500/TOR for the entire system installed, with
gas turbine inlet cooling coil, auxiliary cooling tower, piping, pumps, electrical, controls, foundations,
TOR stands for Ton of Refrigeration, an archaic unit still in common usage. It is the approximate cooling rate that
would produce one ton of ice per day (24 hrs). 1 TOR = 12000 BTU/hr = 3.52 kJ/s (kW).
2
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 M. A. Elmasri, 1990-2008
Chapter 13: Performance Enhancement
engineering and site labour. The installation costs are higher than for electric chillers because of the need
for steam piping and because of the greater heat rejection load.
As shown later, the net thermodynamic impact on a combined cycle’s performance is roughly the same,
whether centrifugal or absorption chillers are used. In view of this, and of the fact that centrifugal chillers
are less costly, easier to install, easier to operate across a range of conditions, and more widely used and
accepted in other applications, they are the system of choice.
Cooling 1 kg/s of air by 10 °C, without condensing any of its humidity, requires a heat removal rate of 10
kW. Cooling humid air to the point of condensing some of its vapour content requires greater heat removal
rates. For instance, cooling 1 kg/s of 100% humid air from 40 °C to 30 °C requires a heat removal rate of 60
kW, six times that for dry air! If the 40 °C air were 80% humid, cooling it by 10 °C would need 37 kW of
heat removal. For 35 °C air at 60% relative humidity, the required heat removal rate for a 10 °C temperature
drop is only 13.5 kW. Thus, the cooling capacity to attain a certain air temperature drop depends strongly
on the humidity.
A typical gas turbine’s output increases by ≈7% per 10 °C of inlet air temperature reduction (§6.3.1.4). If,
under typical circumstances, we use 30 kW of cooling capacity per 10 °C of temperature drop, and given
that a typical modern gas turbine has a specific power on the order of 350 kW per kg/s of airflow, we can
conclude that to gain 25 kW of electrical output from the gas turbine (7% of 350), we have to remove 30 kW
of heat from the air. With an electric chiller COP of 5, this requires expenditure of 6 kW of electricity to
drive the chiller. Thus, under typical circumstances, the gain in gas turbine electric output is roughly four
times the electric input to the chiller. With very humid, hot ambients, the gas turbine power output gain may
be as low as 2½ times the input to the chiller, and with very dry ambients, the gas turbine power output gain
may be as high as ten times the electric input to the chiller.
Figure 13 illustrates the effects of applying inlet air chilling to the design of a large combined cycle for a hot
climate. That example is based on a GE Frame 9FA with a 3-pressure reheat steam bottoming cycle. The
design condition is assumed to be 35 °C (95 °F) ambient at 60% RH. In this example, it is assumed that the
plant design with the chiller will differ from its design without. By cooling the gas turbine inlet air, the
massflow through the engine increases, resulting in higher gas turbine output as well as higher steam
generation in the boiler. However, the exhaust temperature of the gas turbine falls with a cooler inlet. The
boiler needs to be redesigned to cope with the higher mass flow rate of cooler exhaust gas, and still secure
the same throttle temperature at the steam turbine. Thus, it is bigger and costlier, being designed to handle
more gas with the same pinch points and pressure drops. Despite the larger boiler, the increase in steam
generation is not proportional to the increase in gas turbine power and mass flow, due to the decline in the
proportion of the exhaust energy above the pinch.
The x-axis of Figure 13 shows designs for different levels of inlet chilling, expressed as the temperature
reduction of the gas turbine inlet air. Maximum design chilling is by 20 °C (36 °F), which brings the gas
turbine inlet air down to standard ISO temperature of 15 °C/59 °F (but with 100% RH). Whether this is
effected by a centrifugal or absorption chiller is immaterial to the gas turbine gross output, which increases
by about 14%, after accounting for the higher inlet pressure drop due to presence of the finned-tube coil at
gas turbine inlet.
13-13
 M. A. Elmasri, 1990-2008
Chapter 13: Performance Enhancement
Effect of GT Inlet Air Chilling on GT, ST and Gross Plant Output for a Typical 3-P RH
Combined Cycle
0
-9
Output as % of Unchilled Values
115
-18 F
-27
-10
-15
-36
GT
ST, Cen. Chiller
ST, Abs. Chiller
Plant Gross, Cen. Chiller
Plant Gross, Abs. Chiller
110
105
100
95
0
-5
-20
Inlet Air Temperature Change, C
Figure 13. Effect of centrifugal and absorption chilling on gas turbine, steam turbine, and combined cycle gross output
Effect of GT Inlet Air Chilling or Evaporative Cooling on Net Performance of a
Typical 3P RH Combined Cycle
Output & HR as % of Unchilled Values
108
0
-9
-18 F
-27
-36
107
106
Plant Net kW, Cen. Chiller
105
Plant Net kW, Abs. Chiller
104
Plant Net kW, Evap.Cooler
Plant Net HR, Cen. Chiller
103
Plant Net HR, Abs. Chiller
102
Plant Net HR, Evap.Cooler
101
100
99
0
-5
-10
-15
-20
Inlet Air Temperature Change, C
Figure 14. Effect of centrifugal chilling, absorption chilling, and evaporative cooling combined cycle net performance
With a centrifugal chiller, the steam turbine output increases by about 2.5% and the plant gross output by
about 9.5% at the maximum chilling point in Fig. 13.
With an absorption chiller, steam turbine output can decline with chilling, since steam that would otherwise
be available to flow through the steam turbine is consumed by the chiller instead. Thus, at the maximum
chilling point on Fig. 13, steam turbine output is about 4% less and plant gross output is only about 7%
higher than for the unchilled case.
In either case, the gas turbine consumes more fuel due to the increased airflow with chilling. Although the
absorption chiller results in less gross plant output than the centrifugal, the plant net output is about the same
13-14
 M. A. Elmasri, 1990-2008
Chapter 13: Performance Enhancement
for both types. This is because the centrifugal chiller consumes a far greater amount of auxiliary electric
power. Figure 14 shows net performance for the combined cycle. It appears that with the assumptions used,
the centrifugal type has a slight advantage.
The curves of Figs. 13 and 14 change in slope when the extent of chilling results in atmospheric humidity
beginning to condense (at about 7 °C temperature reduction). Once this begins, the chilling load per degree
of incremental air cooling rises due to latent heat. Thus chilling is much more effective if the ambient is hot
but dry; but if that were the case, then evaporative inlet cooling is preferable. Fig. 18 shows the results with
evaporative inlet cooling, which is superior to chilling but limited in its range since it cannot cool the air
below the adiabatic saturation temperature (essentially equal to the ambient wet bulb).
In conclusion, refrigeration systems can be effective in increasing power in both dry and humid
environments, but can be costly as well as detrimental to heat rate. Although overspray fogging can provide
greater benefits at lower cost, refrigeration has the advantage that it works around the gas turbine engine, not
within it, so it is approved by all gas turbine manufacturers, since it does not pose uncertain hazards.
Generally, chilling the inlet of a large combined cycle can "buy" extra hot-day capacity at $400-600 per
additional kW. The lower end of this range applies in dry climates with cooling to above the dewpoint, and
the upper end to humid climates, with the air being cooled to well below the dewpoint. Evaporative cooling
can "buy" extra power at just $200-300 per additional kW of capacity gained, but is limited in its efficacy to
warm and relatively dry ambients, since it cannot cool the air to below the dewpoint.
13.4
SUPERCHARGING THE GAS TURBINE
When a gas turbine is installed at high altitude, it produces less power, but its price does not change, so its
cost per kW increases. One way to restore its output is to supercharge it with a fan. The sort of fan to use is
similar to those used in conventional steam power plants for forced draft of air into the boiler. These are
usually large, slow speed centrifugal fans, reliable, low-tech and relatively inexpensive. The airflow of a
100-MW-nominal gas turbine is comparable with that of a 350-MW-nominal conventional steam power
plant, and a typical forced draft fans provide a pressure rise of 5-10% (50-100 mb or 20-40 “H2O), so will
compensate for 500-1000 m (1650-3300 ft) of site altitude if used to boost pressure before a gas turbine
There are, however, several difficulties with supercharging. First, if the gas turbine’s inlet pressure were
raised but its exhaust pressure were not, the exhaust volumetric flow rate and annulus velocity would
increase, in rough proportion to the percentage boost. Thus, a 10% increase in inlet pressure would increase
exhaust annulus velocity by roughly 10%. This would result in exhaust gas swirling in the exhaust duct, in a
direction counter to the blade rotation. While swirl does occur during normal operation, without
supercharging, as ambients and loads vary, its extent is known and built into the manufacturer’s correction
curves. Should the supercharge introduce excessive levels of it, they may be beyond the manufacturer’s
experience and could lead to unforeseen difficulties, especially in machines with a high exhaust velocity and
high annulus Mach number. Second, the imbalance in pressure between the inlet and exhaust will introduce
additional loads on the thrust bearings. Third, the additional power output may be beyond the generator’s
cooling capacity, which is diminished due to the lower ambient air pressure. Without supercharge, the
diminution of generator cooling capacity is neutralised by the reduction in the machine’s at high altitude.
Due to these difficulties, as well as possible others, supercharging requires additional engineering studies to
be performed by the gas turbine manufacturers to guarantee its performance and to ensure that it will not
adversely affect the machine. Most manufacturers find such studies onerous, especially if they will benefit
very few of their customers. Hence most manufacturers do not approve of this method of boosting output,
although there are many installations built during the 1970’s in which it was used, mostly with
Westinghouse gas turbines.
To understand the thermodynamic impact, we note that a 10% boost in pressure by a fan is accompanied by
an increase in temperature, due to the work of compression. With an exponent of about 0.33 for the
13-15
 M. A. Elmasri, 1990-2008
Chapter 13: Performance Enhancement
relationship between pressure ratio and temperature ratio, corresponding to 86.5% fan polytropic efficiency,
10% boost causes a 3.3% increase in temperature, or about 10 °C (18 °F) temperature rise in an ambient of
15 °C (59 °F). A 10% pressure boost at the inlet, on its own, if it did not change the temperature, would
tend to increase airflow by 10%. Compressor pressure ratio would be unchanged, since its inlet and delivery
pressures would both rise by 10%. Turbine pressure ratio would increase by 10%, since its inlet pressure
would rise 10%, but its exit pressure would be unchanged. The effect would be identical to that of the inlet
pressure drop, discussed in §6.1.4 in detail, except that the ‘pressure drop’ is negative. Engine output would
increase on the order of 15%, as shown in §6.1.4.3. Thus, 10% pressure boost, at constant inlet temperature,
would increase output by about 15%. However, the accompanying increase in inlet temperature, on its own,
would reduce airflow as well output, the latter by roughly 7% (since each 10 °C change in inlet temperature
corresponds to ≈-7% change in power output, as shown in §6.3.1.4). Thus, the effect of a 10% boost by a
fan that is not followed by a cooler is to increase gas turbine output by around 15-7≈8%. If the fan were
followed by a cooler, to cancel its temperature rise, the gain in gas turbine output would be about 15%.
To estimate the net effect, we need to subtract the electric power needed to drive the fan. As shown above,
the air temperature rise in the fan is about 10 °C for a 10% boost starting at ISO temperature. This
corresponds to work done on the air of about 10 kJ/kg, since Cp of air is about 1 kJ/kg-°C. After adding 10
% for the various mechanical and electrical losses, power consumed by such a fan is on the order of 11 kW
per kg/s of airflow. Given that a typical modern gas turbine has a specific power on the order of 350 kW per
kg/s of airflow, we can conclude that for each kg/s of airflow, we can gain 28 kW of electrical output from
the gas turbine if we supercharge it without aftercooling (8% of 350), but we need to expend 11 kW of
electricity to drive the fan. If the fan is followed by a perfect aftercooler, we can gain 53 kW of electrical
output from the gas turbine (15% of 350), for the same expenditure of 11 kW of electricity to drive that fan.
Thus, supercharging can be expected to increase power output from a typical gas turbine by between 2½
times and five times the power consumed by the supercharging fan, depending on whether an aftercooler is
used or not. That increase can even be greater for gas turbines with a high specific power, and may be less
for those with low specific power.
In a combined cycle, supercharging the gas turbine increases the output of the steam bottoming cycle, due to
the larger mass flow rate of the exhaust gas, but decreases its efficiency, due to the lower gas turbine exhaust
temperature. If the supercharging is a retrofit to fixed bottoming cycle equipment, the efficiency reduction
will be greater, since the bottoming cycle equipment will have been already sized for a hotter exhaust at a
lower mass flow rate. If the supercharging is at the design phase, it will lead to larger, more expensive
bottoming cycle equipment, and a smaller decline in bottoming cycle efficiency vis a vis a design without
supercharging.
Figure 15 compares the key performance parameters of three designs of a combined cycle for installation at
an altitude of 1000 m (3280 ft). The base case is unsupercharged, and two designs, one supercharged
without aftercooling and one with a 95%-effective evaporative cooler following the supercharging fan. The
fan in both cases supplies 100 mb (40 “H2O), which represents 11% of the 900 mb ambient. The gain in gas
turbine output is about 9.5% without an aftercooler and almost 17% with the aftercooler. The gain in steam
turbine output is much smaller, due to the decreased exhaust energy, so the gain in combined cycle gross
output is only about 7% without aftercooling and about 12.5% with aftercooling. The fan consumes part of
this gross output, so the gain in net combined cycle output is yet smaller, about 4.5% without aftercooling
and 10% with aftercooling.
13-16
 M. A. Elmasri, 1990-2008
Chapter 13: Performance Enhancement
Figure 15. Effect of design-point supercharging the gas turbine of a dual-pressure combined cycle
The gas turbine heat rate improves with supercharging. The bars showing “GT heat rate” are based on the
conventional definition, fuel input and electricity output, disregarding the fact that the fan also provides
energy to the gas turbine as compressed air (or consumes part of the gas turbine’s electrical output). If the
electricity consumed by the fan were subtracted from that produced by the gas turbine, and a heat rate
calculated on this basis, this too will shows a modest improvement over the unsupercharged gas turbine, as
shown by the bars labelled “(GT+fan) heat rate”. Incorporating the fan without aftercooling creates a
Brayton Cycle with a higher pressure ratio, as though the fan were just another, but less efficient,
compressor stage. Raising the pressure ratio improves Brayton Cycle efficiency, so we still get a little
improvement by adding this less efficient compressor stage. With aftercooling, it becomes reminiscent of a
‘Sprint’ cycle, with evaporative cooling after some compression. This further improves GT+fan cycle
efficiency, but only slightly.
The cooler exhaust reduces bottoming cycle efficiency, so that in spite of the improvement of the (GT+fan)
heat rate, the worsening steam turbine pseudo heat rate overwhelms it, with the result that net combined
cycle heat rate worsens, but only by about 0.25%, with supercharging.
Supercharging is estimated to lower combined cycle specific capital cost, since the power gain is greater
than the increase in cost. In the design cases presented here, the increase in cost includes not only the fan
and evaporative cooler, if used, but also the additional costs associated with the larger, more expensive
bottoming cycle equipment to handle the greater mass flow rate of cooler exhaust gas. The last bars in Fig.
15 show the overall plant specific cost, relative to unsupercharged. On an incremental basis, supercharging
without aftercooling is estimated to cost about $500, and supercharging with aftercooling about $350, both
per incremental net kW of capacity added.
13.5
HEATING THE GAS TURBINE FUEL
Fuel heating has recently come into common practice as a way of improving gas turbine combined cycle
efficiency. Preheating the fuel reduces the amount of fuel needed to achieve a given firing temperature in a
gas turbine, increasing its efficiency, but slightly reducing gross and net plant power output. The reduction
in net power is due to increased auxiliaries as well as a small reduction in gas turbine and steam turbine
gross power outputs. When a gas turbine is fed warmer fuel, it burns less, reducing the mass flow rate and
13-17
 M. A. Elmasri, 1990-2008
Chapter 13: Performance Enhancement
water vapour content of the combustion products, which generate less power as they expand through the
turbine. The boiler also generates a little less steam when it receives a smaller mass flow rate of exhaust gas
with a diminished water vapour content, hence a diminution of steam turbine power.
If fuel heating is effected by recovering heat which otherwise would have been lost, the reduction in power
is minimal. An example would be recovering heat from the flue gases at the stack, lowering stack
temperature, by introducing a water loop that recovers additional heat from the flue gas and transfers it to a
fuel preheater. The additional power loss due to the increased pressure drop on the HRB exhaust gas and
the extra pump power to circulate water through the heat exchangers is small. The gain in efficiency would
well justify the small loss of net output in such cases.
Most combined cycles designed for high efficiency would not have additional heat that could be recovered
at the stack, since the stack may be already at a minimum temperature constraint. Thus, fuel heating would
require utilisation of some heat energy which would otherwise have been used to produce steam turbine
power. The fuel heating source may be steam bled from the steam turbine or from one of the evaporators,
or, preferably, it may be hot water drawn from the boiler economisers. Using steam from the steam turbine
or any boiler that feeds it entails an obvious loss of steam turbine power output. Using water from any
economiser upstream (in the HRB gas path) of a boiler drum reduces the gas energy reaching that drum,
thereby reducing its steam generation. This results in a reduction in steam sent to the steam turbine and a
loss of power. In either of these cases, however, there will still be a gain in overall plant efficiency, since
the energy diverted from producing steam turbine power is of relatively low grade. Converting this lowgrade heat directly to power would be at a lower efficiency than that of the plant as a whole, thus using this
heat to save fuel on a one-to-one basis improves the overall system efficiency.
Fuel heating is qualitatively similar to regenerative feedwater heating in a conventional steam cycle, which
reduces power per unit mass of steam generated, but improves efficiency by saving disproportionately more
fuel.
Figure 16 shows the effect of heating the fuel on a typical triple-pressure reheat combined cycle. In this
example, unheated fuel is initially at 42 °C (108 °F). Heating it up to 130 °C (266 °F) may be accomplished
using water bled from the Low Pressure Economiser (LPE), which absorbs heat from gas downstream of all
evaporators. Thus, the heat transferred to the fuel is solely at the expense of the flue gas temperature at the
stack, and results in improving the heat rate with a negligible loss in net power.
If one wished to further heat the fuel to a temperature higher than permitted by the LPE water exit
temperature, water from the IPE (Intermediate Pressure Economiser) exit can be used. In the example of
Fig. 16, this allows the fuel to be heated up to about 212 °C (414 °F). The extra heat absorbed from the
gases upstream of the LPB result in less LP steam being made and sent to the steam turbine, with an
increasing slope of power loss and a diminishing slope of heat rate improvement.
If one wished to further heat the fuel, HPE exit water would have to be used. In the example of Fig. 16, this
allows heating the fuel up to 298 °C (568 °F). The heat taken from the gas upstream of the IPB reduces IP
steam production. Thus, the loss of power steepens and the gain in heat rate becomes less pronounced.
Considering the declining benefit of raising the fuel to higher temperatures, as well as the potential problems
of handling high-temperature fuel, it is prudent to heat fuel to no more than about 200 °C or 400 °F. This is
the level that can be achieved using IPE exit water in a typical, modern, triple-pressure reheat combined
cycle. It results in a net heat rate gain of about 0.6 % and a net power loss of about 0.3 %.
13-18
 M. A. Elmasri, 1990-2008
Chapter 13: Performance Enhancement
Effect of Fuel Heating by HRB Water on Net Performance of a Typical
Triple-Pressure RH Combined Cycle
Net Output & Heat Rate, % change
0
104
176
248
320 F
392
464
536
608
Net kW change, %
-0.1
Net HR change, %
-0.2
-0.3
Heating Water
from LPE exit
-0.4
-0.5
-0.6
Heating Water
from IPE exit
-0.7
Heating Water
from HPE exit
-0.8
-0.9
40
80
120
160
200
240
280
320
Fuel Temperature, C
Figure 16. Effect of fuel heating on combined cycle net output and heat rate
Modern gas turbines with DLN combustors are sensitive to fuel temperature, due to the delicacy of the premix of the fuel with air and the need to maintain the relationship between flame propagation speed and the
velocities of the pre-mixed air/fuel within the design ranges of the system. Thus, certain of these systems
are designed to operate with heated fuel, and cannot operate across the entire load range without it. Others
are designed to operate with unheated fuel, and cannot operate with heated fuel. Thus, depending on the gas
turbine model, fuel heating may be required or disallowed.
Gas turbines with cooling-air precooling may use energy from the hot compressed air to preheat the fuel.
This wastes exergy, since the hot air from compressor discharge is usually at a temperature higher than
needed for the fuel. This hot air can be best used to generate high or intermediate pressure steam in a
combined cycle, and hot water from the IP economisers is thermodynamically better suited for fuel heating.
In open-cycle gas turbine installations, however, using energy from the cooling-air precooler to heat the fuel
is more efficient than wasting this energy to the atmosphere. In such cases, measures have to be taken to
prevent any possibility of pressurised fuel from leaking into the cooling air. One design is to cool the
cooling-air by passing it through finned tubes outside which fans blow atmospheric air. The air blown by
these fans rises in a duct, and passes over other finned tubes, which carry the fuel. Thus, any minute fuel
leakage would be carried up the warm air plume to the atmosphere, rather than into the gas turbine cooling
air and thence through the hot turbine blades.
Fuel heating to avoid the hydrocarbon dewpoint is discussed in §6.6.1. If the energy source is a fired heater,
it is slightly detrimental to heat rate, and if the energy source is an electric heater it significantly degrades
heat rate. §6.6.1 gives numerical guidelines for the impact on heat rate if fuel is heated by these methods.
13-19
 M. A. Elmasri, 1990-2002
Chapter 14: Steam Turbines
STEAM TURBINES
Revised September, 2002
© Maher Elmasri 1990-2002
14.1
STAGE FUNDAMENTALS
14.1.1 ELEMENTARY VELOCITY TRIANGLES
Co
α2
C1
W1
C1
W1
β1
u
u
Cy1
u
The high-speed steam jet issuing
from the nozzles impinges upon the
buckets, driving them at a
peripheral speed u. The velocity of
the steam relative to the buckets is
W, so that at the inlet to the buckets,
W1 is the vector difference between
C1 and u, and the blades are
Cy2
W2
C2
A typical turbine stage consists of
two blade rows: a row of stationary
nozzles followed by a row of
moving buckets.
The nozzles
reduce the steam static pressure and
enthalpy and increase its kinetic
energy. This requires a converging
passage in subsonic flow. Since no
work is done by the nozzles, the
total energy content of the steam,
i.e. its stagnation (or total) enthalpy,
is unchanged. Thus, in Figure 1, the
nozzle accelerates the steam from
C0 to the much higher velocity C1 at
constant total enthalpy, with the
drop in static enthalpy equal to the
increase in kinetic energy.
Cx
W2
C2
h
ho,1=ho,0
C02/2
h0
Stage Work
C12/2
ho,rr
ho,2
W1 /2
2
h1
W22/2
C22/2
h2
Nozzles
Buckets
Figure 1. Turbine stage fundamentals
14-1
Flow
direction
 M. A. Elmasri, 1990-2002
Chapter 14: Steam Turbines
designed so that W1 is aligned with the blade inlet surfaces. The buckets turn the flow, expelling it at a
relative velocity W2. The extent of convergence, if any, of the flow cross section within the buckets
determines W2 for a given value of W1. The vector sum of W2 and u gives the absolute exit velocity from
the stage, C2.
An imaginary observer moving with the buckets would see them just as a stationary observer would see the
stationary nozzles. Thus, in rotor-relative coordinates, no work appears to be done, and the steam’s rotorrelative stagnation enthalpy does not change. Any drop in static pressure and enthalpy through the buckets
would be balanced by an increase in the relative kinetic energy (W2/2).
14.1.2 STAGE WORK
The work done per unit mass (or power per unit mass flow rate) is the drop in steam total enthalpy, ∆h0.
This can be calculated from the velocity triangles, since power is force times velocity. The tangential force
on the buckets is the rate of change of steam momentum in the tangential direction. The steam absolute
velocity entering the rotor is C1 with a tangential component of Cy1 in the rotor's direction, and the steam
absolute velocity leaving the rotor is C2 with a tangential component Cy2, assumed to be a positive quantity if
its direction is against the rotor's motion, as sketched in Figure 1. The rate of change of tangential
momentum of the steam as it passes through the rotor equals the force exerted on the blades, F,
F = m (Cy1 + Cy2) ..................................................... (1)
where m is the steam mass flow rate. The power P developed by the rotor is the force times the blade
tangential velocity, u,
P = m u (Cy1 + Cy2) .................................................. (2)
and the work done per unit mass of steam, which must equal the drop in steam stagnation enthalpy h0, is
∆h0 = (h01 - h02) = (h00 - h02 ) = u (Cy1 + Cy2) ....... (3)
The drop in stagnation enthalpy across the entire stage is equal to that across the rotor because the stator
does no work, so conservation of energy implies a constant stagnation enthalpy across it.
Introducing the definition of a Stage Loading Coefficient, ψ, which represents the dimensionless specific
work of the stage
ψ = ∆h0 /u2 = Cy1/u + Cy2/u
............................
(4)
i.e. the stage loading coefficient depends on velocity ratios only, which, in turn, depend on the stage
geometry only.
Assuming that the axial throughflow component of velocity, Cx, is uniform allows one to express Cy1/u as
(Cy1/Cx) (Cx/u) and Cy2/u as (Cy2/Cx) (Cx/u). Defining the Flow Coefficient, ϕ, as
ϕ = Cx /u
..........................
(5)
allows one to relate the dimensionless specific work to the geometry as follows:
ψ = ϕ (tan α1 + tan α2) ....... (6)
Since tan α1 = u/Cx + tan β1 = 1/ϕ + tan β1 , then Equation (6) can also be written as
ψ = 1 + ϕ (tan β1 + tan α2)
(7)
Thus we see that stage specific work is proportional to the square of the blade velocity,
∆h0 = ψ u2 ............................
(8)
and that the coefficient of proportionality depends solely on the geometry, i.e. on the blade angles.
14-2
 M. A. Elmasri, 1990-2002
Chapter 14: Steam Turbines
The derivations given above are all based on the fundamental principals of conservation of momentum and
energy, so they apply equally whether the flow through the stage is assumed ideal and frictionless or
whether fluid friction is considered. Fluid friction, will, however, affect the flow velocities used in the
above equations, and will result in a smaller enthalpy drop for a given pressure drop across the stage.
14.1.3 IMPULSE & REACTION STAGES
The buckets may be designed to simply absorb the tangential momentum of the steam jet, without altering
its pressure, in which case the stage is called an Impulse Stage. An impulse stage requires the flow passage
through the buckets to be of uniform cross-section, so that the relative velocity (W) through the passage is
unchanged, and W2=W1. Thus, there is no drop in static pressure and enthalpy across the buckets, so the
wheelspace is at a constant pressure. This avoids the tendency of steam to leak across the blade tip
clearances. It also allows partial arc admission, wherein steam is admitted around only a portion of the
wheel’s circumference, usually at part-load. If there were a pressure difference across the wheel, partial arc
admission would function inefficiently, since the steam would tend to flow circumferentially around the
wheel at its inlet, in an attempt to reach the low pressure space through the rotor sectors which have no
steam admission. Figure 2 illustrates a cross section through an impulse rotor.
C1
W1
u
Figure 2. Impulse buckets
C2
W2
u
Alternatively, the cross-section of the passage through the buckets may be designed to accelerate the relative
flow, in which case the stage is known as a Reaction Stage. A reaction stage requires the flow passage
through the buckets to be shaped like a nozzle, so that the relative velocity (W) through the passage is
accelerated from W1 to a higher value, W2. A reaction stage creates a pressure difference across the bucket
passage.
C1
W1
u
C2
Figure 3. Reaction buckets
W2
u
14-3
 M. A. Elmasri, 1990-2002
Chapter 14: Steam Turbines
The Degree of Reaction, R is defined * as the ratio of static enthalpy drop across the rotor to the stage total
enthalpy drop, i.e.
R = (h1 – h2)/∆h0 ....................................
(9)
which, for the case of uniform axial velocity and equal stage inlet and exit velocities (C2=C0), can be
expressed after some algebraic manipulation as a sole function of the geometry:
R = ½ ϕ (tan β2 – tan β1)} ..........................
(10)
R = ½ {1 + ϕ (tan β2 – tan α1)} ..................
(11)
For an impulse stage with an axial exit (axial C2) and a uniform throughflow velocity Cx, the velocity
triangles appear as shown to the left in Figure 4, with W1 = W2, and β1 = β2. The degree of reaction R is
zero and the stage loading coefficient is ψ = 2.
For a symmetric 50%-Reaction stage with an axial exit (axial C2) and a uniform throughflow velocity Cx, the
velocity triangles appear as shown to the right in Figure 4, with C1 = W2, C2 = W1, α1 = β2, and α2 = β1. The
degree of reaction is R=0.5 and the stage loading coefficient is ψ = 1.
C1
u
W2
C1
u
u
W2
u
Figure 4. Velocity triangles of an impulse stage with ψ=2 (left) and a 50%-reaction stage with ψ=1 (right)
For a given blade speed, u, the impulse stage has twice the enthalpy drop of the 50%-reaction stage, for the
assumptions depicted in Fig. 4. The impulse stage, however, will have a much higher flow velocity issuing
from its nozzle, which increases frictional losses and reduces efficiency. If the velocity of the steam from
the nozzle becomes supersonic, additional losses result from the shock waves which stand in the blade
passages in conjunction with supersonic flow.
14.1.4 PRACTICAL LIMITS ON STAGE VELOCITIES & PRESSURE RATIO
Figure 5 shows the stage enthalpy drop as a function of blade pitch speed for ψ=1 and ψ=2. Most practical
designs will fall between these limits. Since the enthalpy drop per stage is proportional to u2, it is desirable
to have a high blade speed to maximise the work per stage. This minimises the number of stages, which
keeps the size and cost down.
There are, however, two practical limits on maximum blade pitch speed. The first is centrifugal forces
acting on the weight of the blade, tending to pull it out of its root attachment to the wheel, and stressing the
wheel itself. These forces depend not only on pitch speed, but also on tip speed (or pitch speed and blade
length). Since the creep strength of steel depends on temperature, the permissible values are lower for the
higher temperature stages. The second practical limit on blade speed is that high steam velocities cause
greater frictional losses, and if the velocities exceed sonic, also cause shock wave stagnation pressure losses.
Steam velocities are proportional to blade velocity, with the proportionality depending on the velocity
triangles. Impulse stages have a higher ratio of steam velocity to blade velocity. The mark on the curve for
ψ=2 on Figure 5 indicates the point at which the nozzle exit velocity (C1) would reach the speed of sound,
based on a typical steam turbine inlet temperature. This sets maximum blade velocity for subsonic flow at
about 275 m/s (900 ft/s) for an inlet impulse stage. An inlet reaction stage can be run faster, while still
*
This is not the only definition of Degree of Reaction, for example, some base it on pressure drop.
14-4
 M. A. Elmasri, 1990-2002
Chapter 14: Steam Turbines
avoiding sonic conditions. As the steam cools down with expansion to lower pressures, the speed of sound
falls in proportion to the square root of absolute steam temperature, requiring lower blade speeds and/or
higher degrees of reaction to avoid sonic conditions. At the LP stages near the exhaust to the condenser,
supersonic stages may have to be used.
200.0
328
492
656
820
984
1148
1312
Supersonic nozzle exit
@ typical throttle T
180.0
1476 ft/s
86
Stage Enthalpy Drop, kJ/kg
160.0
69
140.0
Impulse
50%-Reaction
120.0
52
100.0
34
80.0
60.0
17
40.0
Practical Range
20.0
0.0
100
150
200
250
300
BTU/lb
350
400
450
Pitch velocity, u (m/s)
Figure 5. Stage enthalpy drop as a function of blade pitch velocity, for an impulse stage with ψ=2 and for a 50%reaction stage with ψ=1
The compromise between the desirability of high blade speed to reduce the number of stages, the desirability
of moderate flow velocities to improve efficiency, and the practical limits set by material strength, all lead to
desirable blade pitch velocities within the approximate range 200-350 m/s (650-1150 ft/s). If the rotational
speed is given, typically synchronous speed for direct generator drive, then the blade speed u implies the
pitch diameter Dp. Thus, for a machine rotating at 3600 RPM (377 radians/s), desirable pitch diameters are
in the approximate range 1.0-1.8 m (40-72 inches). For a machine rotating at 3000 RPM (314 radians/s),
desirable pitch diameters are in the approximate range 1.2-2.15 m (48-86 inches). Towards the lower end of
the desirable speed/diameter range, impulse (or low-reaction) stages are preferred, since they can provide a
higher enthalpy drop for a given blade speed, and towards the higher end of this range, reaction stages are
preferred since impulse stages would have excessive flow velocities at their nozzle exits.
The desirable range of blade speeds/pitch diameters is too narrow for practical designs. At turbine inlet, the
high-pressure, high-density steam typically results in a volumetric flow rate which, in conjunction with the
smallest desirable diameter, would require blades which are too short in proportion to pitch diameter. A
large number of very short blades attached to a large diameter wheel have a large ratio of containing-wall
surface area to flow cross-sectional area, resulting in excessive fluid frictional losses and low efficiency.
These losses are exacerbated if the wheel is unshrouded. Thus, for smaller turbines, it is common practice to
use lower blade pitch velocities and diameters at the turbine inlet, to allow blades of sufficient height, at the
expense of lowering stage work and increasing number of stages. At turbine exhaust, the low-pressure, low-
14-5
 M. A. Elmasri, 1990-2002
Chapter 14: Steam Turbines
density steam typically results in a volumetric flow rate, which, in conjunction with the largest desirable
diameter, would require blades which are too long in proportion to pitch diameter. Such blades, if uniformly
loaded along their length, would require excessive twist and would have a negative degree of reaction at
their hub diameter even if the degree of reaction at their pitch diameter were reasonable. Negative degree of
reaction implies very high nozzle exit velocity followed by compression in the rotor, with high losses and
potential for boundary layer separation. Avoiding this situation requires a non-uniform loading along the
blade length, but this results in less energy being extracted from steam flowing near the hub than from steam
flowing near the tip, increasing secondary flow losses due to the radial gradients in temperature and density.
Thus, for turbines with a large volumetric exhaust flow rate (per LPT path), it is common practice to use
LPT blade pitch velocities and diameters higher than the “desirable range”, to limit (blade height)/(pitch
diameter) ratio, at the expense of greater flow velocities and losses.
The practical range of pitch velocities is therefore 135-480 m/s (440-1575 ft/s). This corresponds to a 3600RPM range of pitch diameters of about 0.7-2.5 m (27½-100 inches), and to a 3000-RPM range of pitch
diameters of about 0.84-3 m (33-120 inches). The high end of this range is only used at the LPT exhaust,
with blades made of Titanium, which has a better strength to weight ratio than steel at typical exhaust
temperatures. For 3000-RPM turbines, the high end of the range is also limited by manufacturing
constraints as well as shipping constraints on the LPT casing diameter. The extremes of the practical range
are not commonly used, and the broad arrows on Figures 5 and 6 indicate the practical range in common use.
2.5
328
2.4
492
656
820
984
1312
1476 ft/s
100 C
50 bar, Impulse
5 bar, Impulse
2.3
1148
250 C
50 bar, Reaction
5 bar, Reaction
2.2
0.5 bar, Reaction
2.1
2.0
Supersonic
nozzle exit
Stage PR
1.9
1.8
1.7
1.6
1.5
1.4
1.3
1.2
1.1
1.0
100
150
200
250
300
350
400
450
Pitch velocity, u (m/s)
Figure 6. Stage pressure ratios as a function of blade pitch velocity, for a typical impulse stage with ψ=2 and for a
typical 50%-reaction stage with ψ=1
The pressure ratio of a stage depends on its enthalpy drop as well as its inlet conditions, pressure and
enthalpy. Figure 6 illustrates the range of stage pressure ratios corresponding to the enthalpy drops given by
Figure 5. An impulse stage, which has its entire pressure ratio across its nozzle, would need a supersonic
nozzle exit if its pressure ratio exceeded 1.83 (the critical pressure ratio for steam). With typical
assumptions, this would correspond to a pitch speed of around 300 m/s if the inlet were at 50 bars, but to a
pitch speed of around 240 m/s if the inlet were at 5 bars. A 50%-reaction stage can run at higher blade
speeds, and achieve higher stage pressure ratios with subsonic flow throughout. For example, with an inlet
14-6
 M. A. Elmasri, 1990-2002
Chapter 14: Steam Turbines
pressure of 5 bars and an inlet temperature of 250 °C, a pressure ratio of about 2.4 and a pitch speed of about
400 m/s can be attained with subsonic flow. With an inlet pressure of 0.5 bars and an inlet temperature of
100 °C, a pressure ratio of about 2.3 and a pitch speed of about 340 m/s can be attained with subsonic flow.
14.1.5 STAGES WITH LOW VOLUMETRIC FLOW
The problem of low volumetric flow usually occurs at the HP inlet stages of medium or small turbines,
resulting in blades which are stubby and short compared to pitch diameter, as illustrated in Figure 7.
Figure 7. Typical HP buckets (courtesy of GE)
To maintain a sufficiently high blade speed without an excessively small ratio of blade height to wheel
diameters, two approaches are used:
(a) Partial arc admission: The wheel diameter is large enough to give a reasonable blade speed at the
desired RPM (at least 0.7 m for 3600 RPM, say), and part of the wheel circumference is blocked, reducing
the axial flow area to less than (π Dp L), where Dp is pitch diameter and L bucket height. This approach is
frequently applied to the first stage only, but rarely to the subsequent stages. Nozzles which admit steam
over only a portion of the wheel’s circumference must be followed by impulse buckets, and these buckets
must be followed by some axial clearance before the nozzles of the next stage, to allow the steam to be
redistributed circumferentially.
(b) Increased RPM: The wheel diameter is reduced to allow full arc admission with buckets of a reasonable
height. The wheel speed is set high enough to produce a satisfactory blade speed, with the smaller wheel
diameter. A gearbox is then necessary to drive the generator at its synchronous speed. This approach
generally results in higher efficiency for smaller turbines.
14.1.6 STAGES WITH HIGH VOLUMETRIC FLOW
The problem of high volumetric flow usually occurs towards the exhaust end of medium or large turbines,
resulting in blades which are slender, twisted, and long in relation to pitch diameter, as shown in Figure 8.
14-7
 M. A. Elmasri, 1990-2002
Chapter 14: Steam Turbines
Figure 8. Typical LP buckets (courtesy of GE)
Figure 9. Double-flow LPT (courtesy of Centrale Costanera)
To avoid excessive blade velocities or blades which are too long in relationship to pitch diameter, two
approaches are used:
(a) Multiple steam paths: The steam flow is expanded through multiple, parallel turbine paths, as shown in
Figure 9.
(b) Reduced RPM: The wheel diameter is increased to provide a large annulus area. The wheel speed is set
at half the generator’s synchronous speed to produce a satisfactory blade speed, with the larger wheel
diameter. A four-pole generator is used rather than a two-pole design to produce electricity at the grid
frequency while turning at the reduced speed. This approach was common in older, large conventional and
nuclear turbines in conjunction with a double-flow LPT, resulting in a so-called “tandem compound”
arrangement. It is uncommon in combined cycles.
14.2
TYPICAL TURBINE CONFIGURATIONS
14.2.1 STRAIGHT CONDENSING, SINGLE CASING
The simplest turbine for power generation is a straight condensing, single casing, single shaft design. The
maximum ratio of exhaust volumetric flow to inlet volumetric flow in such a configuration is about 750.
This is achieved by:
(a) Tripling the pitch diameter, and hence the blade pitch velocity. Thus blade velocity would vary from
135-150 m/s (440-500 ft/s) at the inlet to about 400-480 m/s (1350-1575 ft/s) at the exhaust.
(b) Increasing blade height by a factor of fifty, which increases L/Dp by a factor of 50/3=16.67. The L/Dp
ratio being about .02-.025 at the inlet (with full-arc admission) and about 0.33-0.4 at the exhaust. Partial arc
admission would allow the L/Dp ratio at the inlet to be about 30% higher.
(c) Increasing the axial component of velocity by a factor of five. This would vary from about 60 m/s (200
ft/s) at the inlet to about 300 m/s (1000 ft/s) at the exhaust.
Figure 10 shows illustrates the maximum range of conditions between inlet and exhaust. Naturally, when
such an extreme range is used, the efficiency is less than optimum at either end, with the inlet suffering from
14-8
 M. A. Elmasri, 1990-2002
Chapter 14: Steam Turbines
excessive losses associated with very short blades (in proportion to diameter), and the exhaust suffering
from excessively high velocities and long, twisted blades.
Cx,out ~ 5 x Cx,in
Dp,out ~ 3 x Dp,in
Lb,out ~ 50 x Lb,in
Exit annulus area Aan ~ 150 x inlet
Exit axial velocity ~ 5 x inlet
Exit volumetric flow rate ~ 750 x inlet
Figure 10. Maximum practical range of volumetric flow rates between inlet and exhaust of a straight-condensing, singleshaft steam turbine
The minimum practical volumetric flow rate of steam at the inlet implies a lower bound on the size of
straight condensing turbines running at 3600 RPM of 10-15 MW (14-21 MW for 3000 RPM). Smaller
turbines would have to rotate faster. The maximum practical volumetric flow rate at the exhaust implies an
upper bound on the size of straight condensing turbines of 100-150 MW (140-210 MW for 3000 RPM).
Larger turbines would need a double flow LPT. It should be emphasized that these size guidelines are
approximate, especially the upper bound which depends strongly on condenser pressure. Practical
condenser pressures range between 40 mb (0.58 psia) for an open-loop water-cooled condenser to 160 mb
(2.32 psia) for an air-cooled condenser in a warm climate. Thus, for a given steam mass flow rate, the
exhaust volumetric flow rate has a practical range of 4:1, depending on cooling system and climate.
Figures 11 and 12 show typical, small straight-condensing steam turbines. Figure 13 illustrates the
heat/mass balance of a typical 60-MW-class combined cycle, and the expansion line of its 20-MW-class
steam turbine. The steam mass flow rate at the turbine exhaust is about 20% greater than at its inlet, due to
the addition of IP steam. The steam volumetric flow rate at the turbine exhaust is about 500 times that at its
inlet.
14-9
 M. A. Elmasri, 1990-2002
Chapter 14: Steam Turbines
Figure 11. Typical straight-condensing steam turbine for combined cycle use (courtesy of GE)
Figure 12. Typical straight-condensing steam turbine for combined cycle use (courtesy of Siemens-Westinghouse)
The HP section of such turbines would have 7-10 stages, with a mean enthalpy drop of 45-70 kJ/kg (20-30
BTU/lb) each, and the LP section would have 3-5 stages, with a mean enthalpy drop of 115-160 kJ/kg (5070 BTU/lb) each. Total expansion line enthalpy drop would be on the order of 1200 kJ/kg (500 BTU/lb).
Including the effect of the second steam admission, typical in smaller combined cycles, specific power
output would be on the order of 1100-1320 kW per kg/s (500-600 kW per lb/s) of throttle steam flow rate.
14-10
 M. A. Elmasri, 1990-2002
Chapter 14: Steam Turbines
Gross Power
Net Power
Aux. & Losses
Gross Heat Rate
Net Heat Rate
Gross Electric Eff.
Net Electric Eff.
Fuel LHV Input
Fuel HHV Input
Net Process Heat
59815 kW
58398 kW
1417 kW
6871 BTU/kWh
7037 BTU/kWh
49.66 %
48.49 %
410974 kBTU/h
456022 kBTU/h
0 kBTU/h
Ambient
14.7 P
59 T
60% RH
865 p
950 T
37.79 M
Stop Valve
21520 kW
0.7367 p
91.72 T
45.17 M
100 p
496.8 T
7.475 M
14.7 p
215.5 T
306.7 M
to HRSG
IP
HP
LPB
17.19 p
220 T
0.9537 M
IPB
110.2 p
334.9 T
7.475 M
HPB
922.1 p
534.8 T
37.79 M
292.1 T
280.1 T
444.1 T
358.9 T
871.6 T
558.8 T
LP
CH4
5.305 M
410974 kBTU/h LHV
14.55 p
59 T
301.4 M
15.02 p
1004.6 T
306.7 M
91.62 T
45.73 M
p [psia] T [°F] M [lb/s], Steam Properties: Thermoflow - STQUIK
0 02-21-2002 19:59:12 file=C:\Tflow7\MYFILES\STAG_106B_1996.gtp
GE 6541B
@ 100% load
38295 kW
GT PRO 10.3.2 Maher Elmasri
GT PRO 10.3.2 Maher Elmasri
Net Power 58398 kW
LHV Heat Rate 7037 BTU/kWh
Steam Turbine Expansion Path
1600
865 psia
ft3/s
37.4 lb/s - 35.6
17 kg/s - 1.01 m3/s
1500
1000 F
HP
800 F
1400
1300
ENTHALPY [BTU/lb]
100 psia
218 BTU/lb
506 kJ/kg
HPTL
600 F
Total enthalpy drop
504 BTU/lb
1173 kJ/kg
IP
400 F
1200
287 BTU/lb
667 kJ/kg
1100
.95LPTL
0.7367 psia
.9
.85
1000
Exhaust
.8
45.2 lb/s - 17,770 ft3/s
20.5 kg/s - 503 m3/s
900
800
1.3
1.4
1.5
1.6
1.7
ENTROPY [BTU/lb-R]
1.8
1.9
2
Exhaust mass flow 1.2 x inlet,
volume flow 500 x inlet
Figure 13. Typical heat/mass balance of a small (60-MW-nominal) combined cycle (top) and key parameters of its 20MW-nominal steam turbine (bottom)
14-11
 M. A. Elmasri, 1990-2002
Chapter 14: Steam Turbines
14.2.2 REHEAT STEAM TURBINES
Most combined cycle steam turbines would have a single-flow HP section with an expansion ratio of
between four and five. The steam exhausting from this section is sent to the boiler for reheat, then returns to
the steam turbine’s IP section, in most cases augmented by the additional IP steam generated in the HRSG.
In a typical combined cycle turbine, the IP section would also be of a single-flow configuration, and would
have an expansion ratio of between five and eight. From the IP exhaust the steam would flow to the LP
section, which expands it to the condenser pressure.
In smaller turbines, below roughly 140 MW (for 3600 RPM) – but dependent on condenser pressure – the
LP would be of single-flow design, so that the combined IP/LP resembles a single-casing, straightcondensing, non-reheat turbine. For larger turbines, the LP would be of double-flow (or even triple-flow or
quadruple-flow) design, necessitating a crossover pipe between the IPT exit and the LPT inlet. This
crossover pipe serves as a convenient location for admitting additional (third-pressure) steam from the LP
section of the HRSG. The crossover also serves to separate the HPT and IPT casings, which are usually of
heavy, cast construction, from the LPT casing(s) which are usually of a lighter, cast or fabricated
construction.
Figure 14. Typical 200-MW-nominal reheat steam turbine for a large (550-MW-nominal) combined cycle (courtesy of
GE)
Figure 14 shows a typical configuration of a reheat combined cycle steam turbine in the 200-MW-class, with
a single HPT/IPT casing followed by a crossover leading to a double-flow LPT. Configuring the HP and IP
in a single casing, with opposed flow paths has several advantages. The live HP steam and the hot reheat
steam are both admitted in the middle of this casing and flow towards its two ends. Thus, the temperature
distribution in the casing is smooth, being hottest in the middle and cooler towards both ends. The live HP
steam leaking past the labyrinth seal enters the IP casing, mixing with the hot reheat steam, which is at a
comparable temperature. The forces on the thrust bearings are reduced by the opposed flow arrangement.
Figure 15 illustrates the heat/mass balance of a typical 550-MW-class combined cycle, and the expansion
line of its 200-MW-class steam turbine. The steam mass flow rate at the turbine exhaust is about 35%
greater than at its inlet, due to the addition of IP and LP steam. The steam volumetric flow rate at the
turbine exhaust is about 1200 times that at its inlet.
14-12
 M. A. Elmasri, 1990-2002
Chapter 14: Steam Turbines
Gross Power
Net Power
Aux. & Losses
Gross Heat Rate
Net Heat Rate
Gross Electric Eff.
Net Electric Eff.
Fuel LHV Input
Fuel HHV Input
Net Process Heat
534326 kW
521215 kW
13112 kW
6001 BTU/kWh
6152 BTU/kWh
56.86 %
55.46 %
3206639 kBTU/h
3558130 kBTU/h
0 kBTU/h
Ambient
14.7 P
59 T
60% RH
1800 p
1050 T
229.7 M
Stop Valve
193651 kW
Cold Reheat
0.7367 p
91.72 T
305.7 M
360 p
1050 T
269.5 M
432.8 p
669.6 T
227.4 M
52.91 p
557.3 T
34.44 M
Hot Reheat
14.7 p
202.4 T
1998 M
to HRSG
LP
IP
HP
LPB
60 p
292.7 T
40.72 M
IPB
424.6 p
450.5 T
42.14 M
HPB
1900.3 p
628.6 T
229.7 M
384.3 T
310.7 T
537.8 T
474.5 T
863.2 T
652.6 T
CH4
41.4 M
3206639 kBTU/h LHV
14.55 p
59 T
1956.6 M
15.02 p
1121 T
1998 M
91.63 T
309.4 M
p [psia] T [°F] M [lb/s], Steam Properties: Thermoflow - STQUIK
0 05-23-2002 09:55:51 file=C:\Tflow7\MYFILES\STAG_207FA_1999.gtp
2 x GE 7241FA
@ 100% load
340675 kW
GT PRO 10.3.2 Maher Elmasri
GT PRO 10.3.2 Maher Elmasri
Net Power 521215 kW
LHV Heat Rate 6152 BTU/kWh
Steam Turbine Expansion Path
1700
1199 F
1600
1800 psia
ft3/s
227 lb/s - 108
103 kg/s - 3.06 m3/s
Hot RH
1000 F
HP
1500
432.8 psia
166 BTU/lb
387 kJ/kg
1400
800 F
LPTA1
HPTL
52.91 psia
ENTHALPY [BTU/lb]
600 F
LP
1300
526 BTU/lb
1224 kJ/kg
400 F
1200
LPTL
.95
Total enthalpy drop
692 BTU/lb
1611 kJ/kg
1100
0.7367 psia
.9
Exhaust
.85
1000
306 lb/s - 126,000 ft3/s
139 kg/s - 3,570 m3/s
.8
900
1.3
1.4
1.5
1.6
1.7
1.8
1.9
2
Exhaust mass flow 1.35 x inlet,
volume flow 1167 x inlet
ENTROPY [BTU/lb-R]
Figure 15. Typical heat/mass balance of a large (550-MW-nominal) reheat combined cycle (top) and key parameters of
its 200-MW-nominal steam turbine (bottom)
The HP section of such turbines would have 8-11 stages, with a mean enthalpy drop of 40-55 kJ/kg (18-24
BTU/lb) each, the IP section would have 6-8 stages, with a mean enthalpy drop of 65-90 kJ/kg (28-40
14-13
 M. A. Elmasri, 1990-2002
Chapter 14: Steam Turbines
BTU/lb) each, and the LP section would have 4-7 stages, with a mean enthalpy drop of 115-150 kJ/kg (5065 BTU/lb) each. Total expansion line enthalpy drop would be on the order of 1600 kJ/kg (700 BTU/lb).
Including the effect of the second and third steam admissions, typical in larger reheat combined cycles,
specific power output would be on the order of 1750-2000 kW per kg/s (800-900 kW per lb/s) of throttle
steam flow rate.
14.2.3 COMPARISON BETWEEN STEAM TURBINES FOR CONVENTIONAL PLANTS AND FOR
COMBINED CYCLES
Figure 16. Typical reheat steam turbine for a conventional steam cycle in the 250-MW class (courtesy of GE)
In a conventional steam plant, steam is extracted along the expansion path, whereas in a typical, multipressure combined cycle steam is added along the expansion path.
A small (25-MW-class) conventional steam cycle would have 3-4 feedwater heaters and extract 25-30% of
the throttle steam flow to be condensed in the heaters. Thus, its exhaust steam flow would only be 70-75%
of its throttle flow. By contrast, a combined cycle with a comparably sized steam turbine would likely be of
dual-pressure design, where the IP steam addition would be about 20% of the throttle flow. Thus, its
exhaust flow would be about 120% of its throttle flow.
A 250-MW-class conventional reheat steam cycle would have 6-8 feedwater heaters and extract 30-35% of
the throttle steam flow to be condensed in the heaters. Thus, its exhaust steam flow would only be 65-70%
of its throttle flow. By contrast, a combined cycle with a comparably sized reheat steam turbine would
likely be of triple-pressure design, with an exhaust flow of about 135% of its throttle flow.
Steam turbine specific power, per unit of throttle flow, is therefore much lower for a conventional steam
cycle than for a comparably sized combined cycle.
14-14
 M. A. Elmasri, 1990-2002
Chapter 14: Steam Turbines
Tables 1 and 2 summarise a comparison between steam turbine throttle flow, exhaust flow, and specific
power (per unit of throttle flow) for the two cycle types. Table 1 shows these quantities for a conventional
steam power plant turbine as percentages of the same quantities for a combined cycle turbine; whereas Table
2 shows them for a combined cycle turbine as percentages of the same quantities for a conventional steam
power plant turbine. The top half of the tables is for typical, small (25-MW-class) units, and the bottom half
for larger 250-MW-class units. The three rows in each section allows the user to easily compare the two
types of turbines on the basis of either the same throttle flow (first row), the same exhaust flow (second row)
or the same power output (third row).
Table 1. Throttle flow, exhaust flow, and power output of typical conventional cycle steam turbines relative to
combined cycle steam turbines.
25-MW-class ST
4-heater conventional steam cycle
relative to dual-pressure non-reheat CC
250 MW-class ST
7-heater conventional reheat steam cycle
relative to triple-pressure reheat CC
Relative Throttle
Flow, %
Relative Exhaust
Flow, %
Relative MW, %
100
58
70
172
100
120
142
83
100
100
50
67
200
100
133
150
75
100
Table 2. Throttle flow, exhaust flow, and power output of typical combined cycle steam turbines relative to
conventional cycle steam turbines.
25-MW-class ST
Dual-pressure non-reheat CC relative to
4-heater conventional steam cycle
250 MW-class ST
Triple-pressure reheat CC relative to 7heater conventional reheat steam cycle
Relative Throttle
Flow, %
Relative Exhaust
Flow, %
Relative MW, %
100
172
142
58
100
83
70
120
100
100
200
150
50
100
75
67
133
100
14-15
 M. A. Elmasri, 1990-2002
Chapter 14: Steam Turbines
14.3
STEAM TURBINE EFFICIENCY
A perfectly efficient adiabatic turbine would have an isentropic expansion, resulting in the maximum
attainable enthalpy drop for a given pressure drop. An inefficient turbine would “dissipate” some of the
work it could have generated into heat, raising the steam enthalpy leaving it. Energy is conserved, so that
the work output of an inefficient stage is reduced relative to a more efficient stage that has the same pressure
drop and a lower exit enthalpy.
14.3.1 DESIGN-POINT DRY STAGE EFFICIENCY
The inefficiency of a stage is created by viscous fluid friction. The extent of these frictional losses relative
to the stage’s useful work determine its efficiency. The frictional losses and hence the stage efficiency
depend on the basic, two-dimensional velocity triangles, as well as three-dimensional effects. The threedimensional effects include secondary, transverse, circulating flows superposed upon the primary
throughflow. These secondary flows are created by centrifugal and Coriolis forces acting upon the nonuniform throughflow velocity profiles, which result from fluid friction retarding the boundary layers, both
on the blades and on the hub and tip end-walls.
A stage with a hypothetical, inviscid fluid would be 100% efficient, and the extent of a stage’s inefficiency
depends on the relative magnitude of fluid viscous forces to the useful force exerted by the fluid on the
blades due to its rate of change of momentum. The ratio of momentum forces to viscous forces is
characterised by the Reynolds Number,
Re = ρVL/µ ............................................. (12)
where ρ is the fluid’s density, V the characteristic flow velocity, L the characteristic length of the blade and
µ the fluid’s viscosity. A hypothetical, inviscid fluid would have an infinite Reynolds Number, and the
higher the Reynolds Number the lower the effect of viscosity relative to fluid momentum. This leads to a
“size effect”, causing a larger stage to be more efficient than a geometrically similar smaller stage, if both
have identical velocity triangles and identical steam conditions.
Thus, stage efficiency depends on:
1) Stage geometry, which is defined by the blade angles and shapes and the ratio of blade height to pitch
diameter. Stages with shorter blades relative to the pitch diameter are less efficient.
2) Flow characteristic velocity. Any one velocity, say that of the blade, u, in conjunction with geometry,
determines all the other velocities.
3) Stage size. As given above, larger stages are more efficient.
Modern CFD techniques can be used to predict the efficiency of each stage, but historically, engineers have
relied on empirical methods to compute the efficiency of an entire steam turbine expansion path, including
the efficiency of its constituent “sections” - a section being a group of stages between extraction or
admission ports. The published method of Spencer, Cotton, and Cannon, has been widely used in the
USA [1]. Although the authors are GE engineers, this method has been known to apply just as well to
contemporaneous Westinghouse designs. This empirical method gives the efficiency of the turbine
primarily as a function of its inlet steam volumetric flow rate at a reference throttle pressure and for a
reference first-stage diameter, with correction curves for throttle pressure and first-stage diameter. This is
not surprising, since for given blade angles and RPM, the volumetric flow rate, throttle pressure and firststage diameter are sufficient to determine the three factors affecting stage efficiency described above. Thus,
this method is most accurate when applied to turbines using the set of blade shapes (blade angles) typically
used by GE in each section of a turbine at the time Reference [1] was published.
Spencer, R.C., Cotton, K.C. and Cannon, C.N., "A Method for Predicting the Performance of Steam Turbine
Generators …, 16500 kW and Larger", 1974 revised version of ASME paper no. 62-WA-209.
[1]
14-16
 M. A. Elmasri, 1990-2002
Chapter 14: Steam Turbines
Reference [1] gives numerous other correction factors for efficiency, including factors that account,
indirectly, for steam wetness (see §12.3.2). It also gives procedures for calculating steam leakage rates past
the valve stems and labyrinth seals.
Since Reference [1] was last revised and re-published in 1974, steam turbine aerodynamic design has
improved. Steam turbines procured in the late 1990’s have efficiencies 1% to 3.5% higher than what the
1974 method would predict. Popular heat balance design programs, such as GT PRO use the method of
Reference [1] as a basis, but with modifications and rationalisations to broaden its range of applicability as
well as bring the estimated efficiencies up to date for more modern turbines.
14.3.2 EFFECT OF MOISTURE ON STAGE EFFICIENCY
steam trajectory
moisture trajectory
C1m
W1m
C1
u
W1
Figure 17. Illustration of how moisture droplets retard the buckets by striking them counter to their motion
If the expanding steam becomes wet, the moisture droplets do not follow the steam streamlines because of
their inertia. The larger the droplets, the greater the inability of the steam to make them change velocity and
direction. Thus, as shown in Figure 17, the steam leaves the nozzles at a velocity C1, but the droplets leave
them at a lower velocity C1m, which is closer to the axial direction. Subtracting blade velocity u, one finds
that whereas the relative velocity of the steam entering the buckets, W1, is aligned with the buckets’ flow
passages, the relative velocity of the droplets, W1m, strikes the back of the buckets, retarding them rather
than advancing them. This reduces the stage work, and is usually modelled as a reduction in stage efficiency
using the empirical expression
where:
η = ηdry – β (1-xm) ..................................... (13)
η = corrected stage efficiency (reduced due to moisture)
ηdry = stage efficiency with dry, moisture-free steam
xm = mean equilibrium steam quality within the stage
β = a coefficient, known as the Baumann Coefficient (empirically ≈ 0.65-0.85)
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 M. A. Elmasri, 1990-2002
Chapter 14: Steam Turbines
Due to the speed of the expansion and the accompanying rapid cooling of the steam, the moisture forms at
an equilibrium quality somewhat below one, i.e. the steam becomes supersaturated before liquid droplets
appear. The quality at which liquid droplets form is known as the “Wilson Line” quality, and its value
depends on the speed of the expansion, as well as on the presence of impurities that serve as condensation
nucleation sites. Empirically, the Wilson Line quality is about 0.97, and the moisture correction is applied
only for stages where the mean quality is below this value.
In addition to reducing efficiency, larger moisture droplets may cause blade erosion. Thus, it is common
practice to avoid steam qualities below roughly 0.9, unless special provision is made to mitigate moisture.
Turbines designed to operate with wet steam will have lower flow velocities to avoid erosion (which means
more stages) as well as gutters within the casing from which the condensed liquid is extracted. The liquid
forms a film on the blades and is centrifuged into the gutters. Such designs are common in nuclear-cycle
turbines, which operate with very high moisture content, especially in PWR cycles. Although non-nuclear
turbines have also been built with moisture removal, these designs are uncommon in combined cycles.
Thus, it is common to require a minimum steam quality of 0.88-0.9 at the exhaust of the LPT in combined
cycle designs. This requirement constrains the maximum throttle pressure in non-reheat combined cycles,
particularly if a cool ambient enables a low condenser pressure.
14.3.3 EXPANSION PATH & LEAVING LOSS AT DESIGN POINT
As shown in Figure 1, stage work is the drop in total (stagnation) enthalpy. Total enthalpy is the sum of the
static (true) enthalpy and the kinetic energy of the steam. To the extent that the absolute velocity entering a
stage, C0, and that leaving it, C2, (as shown in Fig. 1) are equal, or nearly equal, the drop in static enthalpy is
equal, or nearly equal, to the drop in total enthalpy. This is the case at the early stages, where velocities
entering and leaving each stage are modest, but not at the later stages. A practical turbine may have a
velocity as low as 60 m/s at the inlet of its first stage, and as high as 300 m/s at the inlet of its last stage
(§12.1.4, §12.2.1). At 60 m/s, kinetic energy is C2/2 = 1800 [m2/s2] = 1.8 kJ/kg (0.77 BTU/lb). At 300 m/s,
kinetic energy is twenty-five times as high, i.e. 45 kJ/kg (19 BTU/lb). Thus, in typical condensing turbines,
interstage kinetic energy is fairly small at the inlet (HP) stages, but significant at the last (LP) stages.
If one were to plot the locus of points entering and leaving each stage, one would obtain an expansion path
on the Mollier (h-s) chart as shown on Figure 18. The dashed line follows the locus of static states and the
solid line the corresponding stagnation states. At inlet to the first turbine nozzle (state 1) the static and
stagnation states are close due to the low velocity. The difference between the static and stagnation
expansion lines increases towards the condenser, due to increasing velocities associated with volume
expansion that exceeds flow-path cross sectional area expansion. The empirical method of Reference [1]
follows the convention used by steam turbine engineers for many years in that it defines stage and group
efficiency based on static enthalpies. It prescribes, in detail, a method for predicting the expansion line
efficiency, as well as plotting it on the h-s chart, all based on static pressures and enthalpies.
The power developed by a group of stages per unit mass flow rate of steam is equal to the difference
between total enthalpy entering and leaving that group, i.e. h01 – h02 in Figure 18:
W = h01 – h02
.......................................... (14)
= (h1+ ½ Cin,12) - (h2+ ½ Cout,22) ...... (14a)
Due to the modest value of Cin,1, the difference between h01 and h1 is usually small. Furthermore, h1 is
usually inferred by measuring the temperature in the steam pipe upstream of the steam turbine. A
thermocouple protruding into the steam pipe measures a temperature much closer to stagnation than static
temperature, since the steam impinging upon the thermocouple well comes to rest, and the steam flowing
past it heats it by friction. Thus, the weak distinction between static and total enthalpy is seldom made at the
inlet to a steam turbine, and the measured value is likely to be closer to total than to static enthalpy. At the
exhaust of a condensing steam turbine, however, the exit velocity at design conditions is usually 200-300
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 M. A. Elmasri, 1990-2002
Chapter 14: Steam Turbines
m/s (600-900 ft/s), which corresponds to
a kinetic energy of 20-45 kJ/kg (8.6-19
BTU/lb). This cannot be neglected, since
it corresponds to anywhere between one
and four percent of the turbine’s work.
This “Leaving Loss” must therefore be
added to the turbine’s exhaust static
enthalpy in order to find its work (or
subtracted from the turbine’s enthalpy
drop):
Specific Enthalpy, h
ho1
h1
Total (stagnation) enthalpies
Static enthalpies
Adjusted for “shell pressure drop”
Actual bleed pressure
W
W = h01 – (h2+ ½ Cout,22) ....... (14b)
D
= h01 – (h2 + LL) ............. (14c)
Pc
= (h01 – h2) - LL ............... (14d)
ho2
EL
ELEP
h2
Specific Entropy , s
Figure 18. Steam turbine expansion line
One drawback in the conventional practice of plotting an expansion path based on static states is that steam
extracted from the turbine comes out at its total, not static, enthalpy. Velocity in extracted steam lines is
usually modest compared to velocities within the turbine, causing static and total states to be relatively close
to one another within the extraction pipes, whereas they may differ substantially within the steam turbine,
especially in the LPT. Thus the intersection between the turbine static expansion path and the static bleed
pressure will indicate an enthalpy lower than that observed in the extraction pipe. An artifice called the
“shell pressure drop” is introduced in the empirical procedure of reference [1] to correct for this discrepancy,
basically by pretending the extraction takes place at a higher pressure, equal to the actual extraction pressure
plus the “shell pressure drop”. No method is given in reference [1] to quantify this “shell pressure drop”, but
its value should vary from about 1-2% for an HP extraction to as much as 50% for the last LPT extraction in
a conventional steam turbine, which is usually at about 0.5 bar (7 psia).
14.4
STEAM TURBINE OFF-DESIGN BEHAVIOUR
14.4.1 NATURAL PROPORTIONALITY BETWEEN PRESSURE AND FLOW
A group of many steam turbine stages of fixed geometry behaves like a choked nozzle of fixed throat area.
At off design, the pressure pi at the inlet is proportional to flow rate and independent of pressure downstream
of the group, i.e.
m vi / pi = C1 An .......................................... (15)
where m is the steam mass flow rate, vi its inlet specific volume, C1 is a constant and An is the throat area of
the equivalent choked nozzle.
If one were to approximate steam by a perfect gas, v=RT/p would lead to
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 M. A. Elmasri, 1990-2002
Chapter 14: Steam Turbines
m Ti / pi = C2 An ............................................. (16)
which illustrates that if the inlet temperature were constant, inlet mass flow rate and pressure would
approximately be in direct linear proportion and inlet volumetric flow rate approximately constant. The
direct proportionality between pressure and mass flow rate is known as “sliding pressure”.
Thus, for a turbine consisting of many stages, operating in sliding pressure, the volumetric flow rate entering
and leaving each stage remains constant at different loads. This results in velocity triangles that remain
invariant at different mass flow rates, both in the magnitude and the direction of all velocity vectors, hence
the efficiency remains essentially invariant with mass flow rate, as long as the pressures are left to float
naturally with mass flow. Where the natural proportionality between pressure and mass flow is disrupted by
an imposition of an artificial pressure, the off-design efficiency may vary significantly from its design-point
value. Such artificial pressure boundaries occur at the condenser, or at a controlled extraction.
14.4.2 EFFECT OF BACK-PRESSURE ON STAGE GROUP INLET PRESSURE
The proportionality between mass flow and pressure at the inlet to a stage group, as described by Esq. (15),
cannot apply when the group ends at an unnatural, controlled pressure. As that controlled exit pressure is
artificially raised at a given mass flow, the inlet pressure must also begin to rise, otherwise the exit pressure
could be raised until it approaches (or exceeds) the inlet pressure !
One formulation that models this phenomenon in a simple way, and which has stood the test of time for 75
years, is due to Stodola (and known as Stodola’s Ellipse), who modified equation (15) as follows:
m vi / pi = C1 An 1− R 2 ................... (16)
where R is the ratio of exit to inlet pressure of the stage group,
R = pe/pi
............................................ (17)
For a "long" group of stages, i.e. one where the exit pressure is much below the inlet pressure, R is a small
number and R2 becomes negligible, so that equation (16) reduces to equation (15). For a "short" group, i.e.
one where the exit pressure is an appreciable fraction of inlet pressure, the group would "unchoke" and
equation (16) would exhibit an appreciable rise in inlet pressure as exit pressure increases (R increases).
14.4.3 EFFICIENCY IN SLIDING PRESSURE OPERATION
From §12.4.2 & §12.4.3, one finds that for a long group of constant-geometry stages, running at constant
speed, with a constant steam inlet temperature:
(1) Mass flow rate and pressure are essentially in direct, linear proportion everywhere.
(2) Volumetric flow rate is essentially the same at any particular location, regardless of mass flow rate.
(3) As a result of 2 above, velocity triangles at any stage are the essentially the same at any massflow rate;
therefore stage efficiency stays constant at part load.
Thus for a typical sliding pressure turbine, where the pressure at inlet is left to float naturally with mass
flow, efficiency stays constant at part load. The only exception being the last few stages before the
condenser where the natural pressure-massflow relationship is disrupted by the "artificial" condenser
pressure. If condenser pressure were somehow controlled to vary in linear proportion to steam massflow,
the velocity vectors would remain the same everywhere at any load and the sliding pressure turbine would
have constant efficiency. The exhaust loss would also remain constant at its design point value. In practice,
however, the condenser pressure is determined by the condenser physical area, the coolant temperature, and
the heat rejection load, and is thus set irrespective of the natural proportionality between flow rate and
pressure prevailing elsewhere in the turbine.
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 M. A. Elmasri, 1990-2002
Chapter 14: Steam Turbines
14.4.4 LEAVING LOSS AT OFF-DESIGN
The term “Leaving Loss” is used to mean the kinetic energy leaving the last stage of the turbine, as
discussed in §12.3.3.
Fig. 19 illustrates typical velocity triangles for the last stage of a condensing turbine. The last stage is
assumed to have been sized so that the absolute exit velocity vector is nearly axial at the design conditions.
u
A
W2
B
.
V
Van =
A an
Design-point
C
D
Leaving Loss
D
C
A
B
Design-point
Annulus velocity (or volumetric flow rate)
Figure 18. Steam turbine expansion line
At off-design, volume flow rate through the stage can vary through changes in massflow and/or condenser
pressure. Reduced volume flow results in absolute exit vectors as illustrated by B and A; whereas increased
volume flow produce vectors as illustrated by C and D. In constructing those vectors, one should note that
velocity relative to the blade remains roughly in the same direction, aligned with the blade's trailing edge,
but varies in magnitude. Addition of the constant blade velocity to this variable-length vector then produces
the absolute exit velocity vectors.
A slight reduction in volume flow produces a vector B that is shorter than design point absolute exit
velocity. One may then expect a reduction in leaving kinetic energy. Large reductions in volume flow rate
will increase the leaving kinetic energy, since in the limit of near-zero volumetric flow, the exit velocity will
approach the blade speed, u, which is usually greater than the design-point steam exit velocity. With volume
flow rates greater than design, the absolute exit velocity increases, which increase the leaving loss.
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 M. A. Elmasri, 1990-2002
Chapter 14: Steam Turbines
14.4.5 EXHAUST LOSS AT OFF-DESIGN
At exhaust steam volumetric flow rates below design value, the distortion to the velocity triangles illustrated
in §12.4.3 (Figure 18) are not only present at the ultimate stage, but may propagate back to the penultimate
stage as well. This results in the steam leaving the penultimate stage at an angle that is mismatched to the
inlet angle of the ultimate stage’s nozzle, causing losses and reducing stage efficiency. At very low
volumetric flow rates, this distortion propagates further back into the turbine, affecting additional stages.
Thus, at reduced volumetric flow rates, a correction to LPT efficiency is needed, over and above the
correction of the leaving loss discussed in §12.4.3.
At exhaust volumetric flow rates above the design value, the last stage’s velocity triangles distortions do not
propagate back into the preceding stages in a significant way, because the flow at the end of the LPT is
usually sonic, or close to sonic, which prevents, or minimises, the ability of distortions to propagate
upstream.
Rather than provide two corrections, one for changes in LPT efficiency, and another for leaving loss, it is
customary to pretend that LPT efficiency remains invariant, and to provide one “Exhaust Loss” correction,
which combines the two effects.
Typical Exhaust Loss Curve - 23"x65" LS Bucket @ 3600 RPM
50
45
Exhaust Loss, BTU/lb
40
35
30
25
SCC
TF Model
20
K.E.
15
10
5
0
0
200
400
600
800
1000
1200
1400
Annulus Velocity, ft/s
Figure 19. Typical “Exhaust Loss” curve, compared with exit kinetic energy, or “Leaving Loss”
Figure 19 shows a typical exhaust loss curve along with a curve labelled "K.E." that only shows leaving loss
(exit kinetic energy). The exhaust loss is approximately equal to the leaving loss at the design point as well
as at volume flow rates greater than design. At low volume flows, however, the exhaust loss substantially
exceeds the leaving loss. This is because the empirical exhaust loss correlation embodies a correction for
declining efficiency of the last stages at off-design flow rates. This efficiency decline is more pronounced at
low volume flow rates, where the flow disturbance propagates several stages back into the machine.
Another effect lumped into the exhaust loss curve is that of the turbine exhaust diffuser. In theory, a perfect
diffuser would be able to recover the exit kinetic energy, using it to create a static pressure behind the last
stage much lower than the condenser pressure. In practice, efficient diffusers are hard to design in the tight
geometry of the exhaust hood, particularly in downdraft exhausts, but some pressure recovery is
accomplished, especially in axial exhausts. Rather than attempt to compute this effect separately, it is
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 M. A. Elmasri, 1990-2002
Chapter 14: Steam Turbines
customary to assume that the pressure behind the last stage is equal to condenser pressure, and to include the
beneficial effect of the diffuser into the exhaust loss curve as well.
Thus, the exhaust loss curve is not a single, well-defined physical quantity, such as exit kinetic energy
(leaving loss), but corrects for a plethora of effects, of which leaving loss is but one. The exhaust loss curve
is an empirical correlation for an enthalpy difference, called exhaust loss (EL), which when added to the
enthalpy at the expansion line end point (ELEP), allows one to correctly compute turbine work per unit mass
from
W = h1 – (hELEP + EL) ......... (18)
where, by construction, ELEP is found by a self-consistent procedure to find the expansion line efficiency,
such as Reference [1], and where said expansion line is assumed to end at the condenser pressure, and its
efficiency is assumed invariant with flow rate.
14.4.6 PRESSURE CONTROLS & THEIR EFFECT ON EFFICIENCY
Left alone, the natural proportionality of pressure to flow rate results in an invariant volumetric flow rate,
hence in steam turbine off-design efficiency that is essentially unchanged from its design-point efficiency.
However, allowing the pressure to slide at reduced steam mass flow rate is not always convenient from other
aspects of power plant operation. For instance, operating a heat recovery boiler at pressures substantially
below the design-point pressure results in steaming within its economisers. A HRSG drum sized for a
certain pressure may also be too small to deal with volumetric swell during transients at reduced pressures.
Thus, in combined cycles, the throttle pressure may be allowed to slide to a certain extent, down to 75% of
design-point pressure, say, but no further. Some pressure control mechanism, typically a valve or a series of
valves, must then be used at the turbine inlet, to maintain the pressure on the HRSG-side at an appropriate
value when steam turbine inlet mass flow rate is significantly reduced.
Another need for pressure controls arises when steam is extracted from the turbine for cogeneration. If the
“flow to following stage”, i.e. the flow continuing past the extraction port, is reduced, the pressure at the port
will fall proportionately, unless some pressure control device were interposed between the extraction port
and the following stage. The flow to following stage may be reduced either by extracting more steam, or by
any factor that reduces upstream flow.
Whenever a device is used to artificially control pressure in a steam turbine, contradicting the natural sliding
pressure proportionality between pressure and flow, losses are incurred. These are discussed below.
14.4.6.1 Group with controlled exit pressure
A group of stages where the inlet pressure slides in concert with massflow but where the exit pressure is
artificially fixed may occur upstream of an auto-extraction or of the condenser. Such a group will be subject
to an efficiency penalty at its downstream stages at off-design, as described for the exhaust loss above.
14.4.6.2 Group with throttle-controlled inlet pressure
A throttle valve at the inlet to a group can be used to maintain the upstream pressure, before the valve. Once
the steam passes through the valve, it must "fill" the constant area nozzles it encounters. Thus if one wishes
to maintain pressure upstream of the valve at its design point value but at half the design point mass flow,
the valve will be closed such as to throttle the flow to a reduced pressure so as to double its specific volume.
Fig. 20 illustrates the situation. With valve wide open, points 1 & 2, before and after the valve, would
coincide if one neglects VWO pressure drop. To maintain the upstream pressure P1 as massflow is reduced,
the valve would have to progressively close, throttling from State 1 to States 2a and 2b as the mass flow
declines. States 2a and 2b have to be found by the constraints that (i) they possess the same enthalpy as
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 M. A. Elmasri, 1990-2002
Chapter 14: Steam Turbines
State 1 and (ii) they possess combinations of pressure and specific volume that satisfy equation (16) at the
stage-group exit pressure, P3.
With throttle control, the expansion lines 2-3, 2a-3a and 2b-3b all have the same efficiency, since they
correspond to a sliding pressure at 2, i.e. blading "internal efficiency", excluding the throttling process,
remains at its design-point value. It is obvious, however, that throttling reduces the enthalpy drop to a given
back pressure, i.e. the expansion 2b-3b produces less specific work than 2a-3a, which produces less than 2-3.
A definition of "overall efficiency" is sometimes used to lump together valve throttling loss with internal
expansion line efficiency. This would be defined by path 1-3 at full flow and by hypothetical paths 1-3a and
1-3b at lower flows, which would represent a steep decline in apparent efficiency as flow is reduced.
Throttle control therefore preserves internal efficiency but significantly degrades overall efficiency. Due to
its simplicity, however, it is frequently used both at live steam admission and at auto-extractions.
Inlet throttle valve
Specific Enthalpy, h
1
Auto-extraction valve
1
2
3
1
1
2
2a
2
3
2b
1-2-3 : Valve Wide Open
Pc
3b
1-2a-3a and 1-2b-3b show
states as valve is closed to hold
P1 constant at lower flows
EL
3a
3
Specific Entropy , s
Figure 20. Throttle Control with a single valve results in a throttling followed by expansion at the nominal efficiency
14.4.6.3 Group with nozzle-controlled inlet pressure
With this method upstream pressure is controlled by varying the flow area to the first set of blading. In
certain practical constructions, mostly for auto-extractions, the flow area through the inlet nozzle is varied
by swiveling the nozzle blades, much like IGV's in a compressor. A more-common construction is to vary
the arc of admission by having a number of nozzle chambers feeding nozzle sectors, each being
independently controlled by its own throttle valve. This is illustrated by Figure 21.
At full flow, all valves are wide open and states 1A, 1B and 1C are coincident. If one neglects wide-open
valve pressure drop, they would also coincide with state 1.
At slightly reduced flow P1 can be maintained by partially closing one of the valves, C in Fig. 21. Thus the
flow through sectors A and B proceeds without throttling (other than by the small pressure drop
corresponding to the wide-open valves), whereas the flow through sector C is first throttled from 1 to 1C to
expand in volume and fill its nozzle sector area at reduced massflow. The flow through the wide-open
sectors expands through the governing stage from 1A, 1B to 2A, 2B; whereas the flow through the partially
closed sector does less work per unit mass as it expands from 1C to 2C. The average enthalpy entering the
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 M. A. Elmasri, 1990-2002
Chapter 14: Steam Turbines
subsequent group stages is represented by state 2. The subsequent group stages behave as for sliding
pressure, i.e. P2 and h2 can be determined by iteration according to equation (16) using the known value of
P3 and the fixed area of the first nozzle of the group.
2A, 2B
A
B
3
ROTOR
1A, 1B
1C
C
h
P1
1
P2
1'
2A, 2B
Valves A, B are wide-open
Valve C is partly shut
2C
2
P3
3
s
Figure 21. Model of a Multi-Valve Control system with a variable arc of admission
At reduced flow, P2 will fall, increasing the pressure ratio across the governing stage. Thus the nozzle
sectors with wide open valves will operate through a pressure ratio (1A,B):(2A,B) greater than design.
Steam will issue from those nozzles and impinge on the blades at a velocity greater than at design. The
nozzle sector with partially closed valve will operate through a pressure ratio 1C:2C, below design. Steam
issuing from that nozzle group will be at a velocity lower than at the design-point. The efficiency of the
governing stage will therefore be impaired. The efficiency of the group stages, 2-3 is almost equal to its
design point value, since they are effectively in sliding pressure, but still somewhat suffering from the
asymmetry of flow conditions entering them from the governing stage.
Greater reduction in flow, beyond the condition illustrated in Fig. 21, would eventually result in valve C
being fully closed. Further reduction would be accompanied by closure of the next valve, B for instance,
etc... With such sequential operation of valves, the effective pressure drop for the valve system as a whole,
illustrated by 1-1' on Fig. 21, would go through loops as shown on Figure 32.
In many practical situations, it is advantageous not to operate the valves sequentially but to overlap them,
particularly at very low flows. This is because as P2 falls, a fully-open sector with the others fully closed
would experience a very large pressure ratio, leading to very high velocity steam issuing from its nozzles.
This leads to inefficiency as the velocity ratio (the ratio of steam tangential velocity to blade peripheral
velocity) varies greatly from design, as well as excessive forces on the rotor blades, since steam at
abnormally high velocity impacts them. The periodicity of these excessive forces, as blades pass through
the high-velocity steam jet once per revolution, can create dangerous fatigue. Ideally, therefore, the valves
would operate sequentially at flow rates slightly below design and would overlap to an increasing extent at
flows greatly below design, approximating simple Throttle Control at very low flows. The sequencing of
the valves and their overlap is implemented by the shapes of the cams which push the valve stems to actuate
them. These cams are mounted on a camshaft, rotated by a hydraulic crank to close or open the valve
system to raise or lower the pressure, as shown in Figure 22.
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 M. A. Elmasri, 1990-2002
Chapter 14: Steam Turbines
Figure 22. Steam chest of a multi-valve controlled steam turbine inlet, showing the cams used to actuate the valves
(courtesy of GE)
Figure 23. Cut-away view of an auto-extraction steam turbine for cogeneration (courtesy of GE)
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 M. A. Elmasri, 1990-2008
Chapter 15: Condensers & Cooling Towers
CONDENSERS & COOLING TOWERS
Revised September, 2008
© Maher Elmasri 1990-2008
Power plants reject a large amount of heat to condense the steam exhausted from the steam turbine. In a
conventional steam power plant, without cogeneration, this heat rejection rate varies from almost twice the
gross power output for a small (≈25 MW), non-reheat conventional steam plant, to being about equal to the
gross power output for a large (≈500 MW) reheat conventional steam power plant. For GTCC plant, the
steam turbine represents roughly one-third of the net power output, so the total heat rejection at the
condenser is smaller in proportion to plant gross power output. It varies from about 80% of CC power
output for a small (≈25 MW) plant with a non-reheat steam cycle, to about 60% of CC power output for a
large (≈500 MW) plant with a reheat steam cycle. Although the heat rejection as a percentage of plant net
output is lower for GTCC plant than for conventional steam plant, the heat rejection per unit of steam
turbine output is much greater in GTCC plant, because they add steam to the turbine at lower pressures
instead of extracting it for feedwater heating. Heat rejection of GTCC plant varies from about 250% of
steam turbine power output for a small (≈25 MW) plant with a non-reheat, dual-pressure steam cycle, to
about 165% of ST power output for a large (≈500 MW) plant with a triple-pressure, reheat steam cycle.
Table 1 may be used to make an approximate estimate of the heat rejection magnitude.
Table 1. Heat rejection magnitude for typical power plants of various types & size
Conventional Steam Plant
Gas Turbine Combined Cycle Plant
Qrej as % of plant output
Qrej as % of plant output
Qrej as % of ST output
Small (≈25 MW) plant
200
80
250
Large (≈500 MW) plant
100
60
165
The many types of cooling systems used in power plants are summarised in the chart below and described in
this chapter, except that cooling systems which are very uncommon are not described in any detail.
Condensers
Water-Cooled Surface Condenser
CW inside tubes, condensation on outside surface
Open Loop
CW from ocean, lake, river
Cooling Tower
Cools water in closed loop
Mechanical Draft
Wet
Air-Cooled Condenser
CA outside tubes, condensation on inside surface
Wet-Dry
Dry Mechanical Draft
Natural Draft
Dry
Wet
Dry
15-1
Dry Natural Draft
Wet Surface
(mechanical draft)
 M. A. Elmasri, 1990-2008
Chapter 15: Condensers & Cooling Towers
15.1
WATER-COOLED SURFACE CONDENSERS
The most common type of steam condenser consists of a large number of horizontal tubes within which
flows cooling water. The steam exhausting from the steam turbine flows over the exterior of these tubes,
condensing on their cool surface. A shell confines the condensing steam, which is at a pressure below
atmospheric.
Condenser tubes were commonly made of brass for many years, but recent trends are to use stainless steel
tubes with fresh cooling water and titanium tubes with seawater Most condensers use tubes with diameters
in the range 19-38 mm (¾”-1½”). Typically they are of single-pass or two-pass design, and the number of
tubes is determined so as to correspond to a reasonable velocity for the cooling water flowing within.
Ideally, this velocity should be in the range 1.5-3 m/s (5-10 ft/s). Lower velocities result in a low internal
heat transfer coefficient, and higher velocities result in a high pressure drop, as well as possible tube erosion
since the cooling water usually contains some dirt particles.
Condenser shells are usually made of carbon steel with stiffening ribs. The shell skin must be thick enough
to withstand the vacuum forces and for structural integrity. The shell also supports numerous tube sheets,
also made of carbon steel, which hold the tubes in place, at intervals along their length, to prevent them from
vibrating. These tube sheets also help to stiffen the shell against the vacuum forces that tend to collapse it.
15.1.1 OPEN-LOOP WATER-COOLED SYSTEM
Open-loop systems utilise a natural source of water, such as the ocean, a lake, or a river. This means that the
allowable temperature rise, generally referred to as range, must be low, to avoid environmental damage to
aquatic flora and fauna in the vicinity of the plant. Environmental permits are likely to restrict this
temperature rise to a very low value, such as 4-5 °C (7-9 °F). Since condensing a kg of exhaust steam
requires heat removal of over ~ 2000 kJ, and heating a kg of water by ~ 4-5 °C absorbs ~ 20 kJ, the cooling
water flow rate in an open-loop system is about one hundred times the steam turbine exhaust flow rate. By
contrast, condensers designed for cooling tower applications would normally have a range of 10-20 °C (1836 °F), and would therefore handle cooling water flow rates 30-60 times the steam turbine exhaust flow rate.
Water available from a natural source is usually much cooler than water that can be obtained from a cooling
tower at the same location. This, along with the low range permitted for the natural source, may result in a
thermodynamically achievable condenser pressure lower than economically optimum, in view of the cost
penalty of a large steam turbine exhaust-end. There is less impetus to squeeze the pinch temperature
difference between the condensing steam and the warm CW exiting the condenser as in most cooling tower
applications, and so in many cases the required condenser surface area is smaller than in typical cooling
tower applications.
Thus, open loop condensers tend to have higher CW flow rates and, in some cases, lower surface areas than
their cooling tower counterparts. Hence, they tend to employ tubes with larger diameters, tend to favour a
single-pass arrangement, and tend towards a larger number of shorter tubes than equivalent condensers
designed for cooling tower applications.
Open loop condensers, with their low range and large CW flow rates demand much larger circulating water
pumps and pipes than systems with a cooling tower. Also, an open-loop system frequently entails long CW
pipes or tunnels, to reach a sufficient depth of water in the natural source and to separate the intake and
discharge by an adequate distance, to prevent the warm discharge from heating up the intake. The longer,
larger piping, along with the intake and discharge concrete structures, result in piping and civil costs that
may equal or exceed the cost of an alternative cooling tower system.
For a typical fresh water source, the three main components of an open loop system, condenser, pumps, and
piping/civil would cost roughly $6 million, installed, for a 100-MW-nominal steam turbine, or about 60
$/kW of steam turbine output, compared with about $250/kW for the steam turbine itself.
15-2
 M. A. Elmasri, 1990-2008
Chapter 15: Condensers & Cooling Towers
With seawater, Titanium tubes are favoured due to its resistance to galvanic corrosion. A Titanium-tubed
condenser would cost about 40% more than a stainless-steel-tubed condenser. The biggest site-specific
variable for the cost of the seawater cooling system is the piping and civil component, which is likely to be
higher than with a river or lake.
15.1.2 WATER-COOLED CONDENSER WITH MECHANICAL DRAFT COOLING TOWER
This is the most common cooling system in combined cycle plants. The discussions in §15.1.1 all apply,
except that now the CW range and condenser design is not governed by environmental constraints as much
as by the combined techno/economics optimisation of the whole system of condenser, cooling tower, CW
pumps and pipes. These generally lead to a CW range of 10-20 °C (18-36 °F), and to a CW flow rate 30-60
times the exhaust steam.
Figure 1 shows overall dimensions and proportions of a condenser for a 100-MW-nominal steam turbine,
designed for use in conjunction with a cooling tower. The thermodynamic parameters of this design are
illustrated in Figure 2. Such a condenser contains about 10,000 stainless steel tubes, 22 mm ( 7/8” ) in
diameter, and about 10 m (33 ft) long, for a total condensing surface area of about 7,000 m2 (75,000 ft2). It
should cost about $1.5 million (~$200/m2 ~ $20/ft2) and weigh about 135 tonnes (~300,000 lbs). Cost per
unit weight is ~ $11/kg ~ $5/lb (similar to a medium HRSG). These costs are equipment only, with
additional costs for shipping and installation.
The above rules of thumb on condenser surface area should be used in context, since surface area depends
very strongly on the “pinch” temperature difference between the warm CW leaving the condenser and the
steam saturation temperature. This is shown as “CW approach to condensate” in Fig. 2, and its value in our
example is 2.4 °C (4.3 °F). Such a small value is typical in systems with cooling towers, where the
economics favour a large CW range to minimise CW flow rate, and thus the sizes of the CW piping and
pumps. For condensers using an open loop CW source, such as rivers or the ocean, this pinch is usually
much larger (and the condenser area much smaller), since the CW range is usually very small to minimise
environmental impact.
An instructive comparison between a typical Heat Recovery Boiler and a typical Surface Condenser, both
for a reheat combined cycle with a 100-MW-nominal steam turbine, is shown in Table 2. Typically, the
surface condenser cost is about 1/12th the boiler’s, and its heat transfer surface area is about 1/20th the
boiler’s. With forced convection of water on the inside of the tubes and condensation of steam on the
outside, the overall heat transfer coefficient is very high in a surface condenser, almost one hundred times as
high as the average overall coefficient for a heat recovery boiler.
Table 2. Comparison of key overall parameters for Heat Recovery Boilers and Condensers, for a typical reheat
combined cycle with a 100-MW-nominal steam turbine.
Heat Transfer Duty, MW
Approximate Cost (equipment only)
LMTD (averaged for HRSG), °C
Overall heat transfer coefficient (averaged for
HRSG), W/m2-°C
Surface Area, m2
LMTD (averaged for HRSG), °F
Overall heat transfer coefficient (averaged for
HRSG), BTU/hr-ft2-°F
Surface Area, ft2
Heat Recovery
Boiler
Air-cooled
Condenser
Water-cooled
Surface
Condenser
260
$ 18 m
40
160
$ 18 m
16
160
$ 1.5 m
7
40
50
3,500
165,000
72
200,000
29
7,000
12
7
9
600
1,800,000
2,150,000
75,000
15-3
 M. A. Elmasri, 1990-2008
Chapter 15: Condensers & Cooling Towers
1
2
FOR QUALITATIVE INDICATION ONLY
3
4
5
6
7
8
Two-pass, 10,000 SS tubes, 10 m (33 ft) long, 22 mm (7/8") dia.
Surface area ≈ 7,000 m2 (75,000 ft2)
Weight ≈ 135 tonnes
Cost (equipment only) ≈ $ 1.5 million
A
A
B
B
E
C
F
A
B
C
D
E
C
D
Typical Practice:
Tube diameters ≈ 19-38 mm (3/4" - 11/2")
One or two passes, water velocity ≈ 1.5-3 m/s (5-10 ft/s)
Condenser cost (Stainless) ≈ $200/m2 ($20/ft2)
Condenser cost (Stainless) ≈ $11/kg ($5/lb)
Titanium tubes with seawater, cost +40%
D
E
Thermoflow, Inc.
Company: Thermoflow
User: Maher Elmasri
WATER-COOLED CONDENSER
F
A
B
2.8 m
7.4 m
C
D
10.5 m
1
1.3 m
E
F
G
H
I
J
4m
3.2 m
-
-
-
-
2
3
4
5
PLAN
Date: 07/22/02
F
Drawing No:
c:\Tflow8\MYFILES\GTPRO.GTP
6
7
8
PEACE/GT PRO 10.9
1
2
FOR QUALITATIVE INDICATION ONLY
3
4
5
6
7
8
A
A
A
STEAM INLET
B
E
B
TRANSITION PIECE
C
CIRC. WATER INLET
G
F
H
HOTWELL
D
D
C
CONDENSATE OUTLET
CIRC. WATER OUTLET
C
D
B
E
E
Thermoflow, Inc.
Company: Thermoflow
User: Maher Elmasri
WATER-COOLED CONDENSER
F
A
B
13.4 m
1.9 m
1
C
D
10.5 m
0.97 m
2
E
F
2.0 m
4.5 m
3
G
3.2 m
H
0.51 m
I
J
-
-
4
5
SIDE ELEVATION
Date: 07/22/02
F
Drawing No:
c:\Tflow8\MYFILES\GTPRO.GTP
6
7
8
PEACE/GT PRO 10.9
Figure 1. Water-cooled condenser for a 100-MW-class steam turbine
15-4
 M. A. Elmasri, 1990-2008
Chapter 15: Condensers & Cooling Towers
Exhaust steam
71.1 M
26.83 TWB
100% RH
3164 M
26.83 T
Typical Ratios
L/S ≈ 30-60
L/G ≈ 1-2
M/S ≈ 0.8-1.2
“S”
0.0483 P
32.27 T
0.0754 M
Fans 755 kW
Condenser
Hot CW
3195 M
“L”
Air
10.83 TWB
60% RH
3114 M
Cooling Tower
29.83 T
“G”
Cold CW
3195 M
17.83 T
15 T
CW Pumps
919 kW
“M”
to HRSG
71.17 M
Condensate
pump 56 kW
35
0.0483 bar
TEMPERATURE [C]
Exhaust Steam
32.27
∆Ta = CW approach to condensate, say ~ ITD/6
30
ITD
Typically
20-30 °C
(36-54 °F)
Blowdown
12.45 M
Makeup
62.25 M
29.83
26.83
25
Cooling Water
∆Tb = CW Range
say ~ ITD/2
Air wet-bulb
temperature
20
17.83
15
∆Tc = CW approach to wet bulb
say ~ ITD/3
Condenser 160514 kW
CT 160514 kW
10.83
10
0
20
40
60
80
100
120
HEAT TRANSFER [.001 X kW]
140
160
180
200
Figures 2 (top) and 3 (bottom). Flow rates and thermodynamic states for a water-cooled condenser and its cooling
tower, sized for a typical 100-MW-class reheat combined cycle steam turbine under ISO ambient conditions. Some
typical ratios & rules of thumb of key parameters are indicated on the diagrams
15-5
 M. A. Elmasri, 1990-2008
Chapter 15: Condensers & Cooling Towers
15.1.3 MECHANICAL DRAFT WET COOLING TOWERS
Figure 4. Mechanical draft wet cooling towers (courtesy of Marley Co.)
This is the most commonly used cooling system in combined cycles. A portion of the cooling water is
evaporated in an air current in the cooling tower, drawing its latent heat from the rest of the water, thereby
cooling it. Since the air current also warms up, some heat rejection in the form of sensible heat to the air
augments the latent heat drawn by the evaporated water.
In a typical cooling tower, the ratio of airflow to CW flow (the “L/G” or “liquid-to-gas ratio”) is between
one (for a cool ambient) and two (for a hot ambient); and the sensible heat absorbed by the air is 10%-25%
of the total heat rejection duty. This leaves 75%-90% of the heat rejection duty to be carried by the latent
heat of the evaporating water within the tower. Since the condensing steam is usually at a quality between
80% and 90%, and since the latent heat of condensation of the steam within the condenser is nearly equal to
the latent heat of evaporation of the water within the tower, the water evaporation rate is 60% -80% of the
exhaust steam flow rate.
In addition to evaporation, some water is lost by drift (water lost as liquid droplets entrained with the air).
Additional water is lost by blowdown, to rid the tower of concentrated contaminants left over after
evaporation. Typically, drift and blowdown are very small as percentages of the circulating water, their
combination being about ½%, but since the evaporated water is itself quite small, typically 1½% - 2½% of
the circulating water, drift and blowdown combined are usually 20-35% of the evaporated water. After
considering these effects, one finds that typical tower water consumption is 85%-105% of the condensing
steam flow. This leads to a common rule of thumb that cooling tower makeup is approximately equal to the
condensing steam flow. This rule is slightly conservative in most practical cases.
15.1.4 ALLOCATION OF THE COOLING SYSTEM’S AVAILABLE TEMPERATURE DIFFERENCE
In a wet cooling tower, the ultimate heat sink is the ambient air wet-bulb temperature. The total available
temperature difference between the steam saturation temperature and the ultimate heat sink, is called the
Initial Temperature Difference (ITD), and is shown in Figure 3.
In a wet cooling tower, this ITD is usually in the range 20-30 °C (36-54 °F). It has to be allocated to three
components, as shown in Fig. 3:
ITD = ∆Ta + ∆Tb + ∆Tc
where
15-6
............................ (1)
 M. A. Elmasri, 1990-2008
Chapter 15: Condensers & Cooling Towers
∆Ta is the CW approach to the condensate in the condenser (the condenser “pinch”)
∆Tb is the CW range
∆Tc is the CW approach to the ambient wet bulb, as it leaves the cooling tower.
Reducing ∆Ta requires a larger condenser surface area and cost. Reducing ∆Tb requires a larger CW flow
rate and hence larger, more expensive pumps and pipes, as well as greater pump power consumption.
Reducing ∆Tc requires a larger contact area between the CW and the ambient air in the tower, increasing the
cost of the cooling tower. In designing the system, the optimum economic allocation of the ITD will be sitespecific and should be done on a case-by-case basis, since it depends on many parameters. In the absence of
detailed analysis, a reasonable rule of thumb is to allocate about 1/6th of the ITD to ∆Ta, 1/2 to ∆Tb, and
1/3rd to ∆Tc.
15.1.5 COOLING SYSTEM COST, CONSUMPTION OF AUXILIARY POWER & WATER
With a cooling tower the CW flow rate and hence the cost of CW pumps and piping is less than with an
open-loop system, but there is the additional cost of a cooling tower. For a typical system, the four main
components: condenser, cooling tower, pumps and piping/civil would cost roughly $7 million for a 100MW-nominal reheat combined cycle steam turbine, or about 70 $/kW of steam turbine output. This
compares with about $250/kW for the steam turbine/generator itself.
The cooling tower fans have to pump an air mass flow rate roughly equal to the CW mass flow rate, about
40-50 times the condensing steam mass flow rate. Typical fan pressure is about 1.25 millibars (½ “ H2O).
The CW pumps have to overcome the friction pressure drops in the piping and within the condenser, and
also have to elevate the water to the upper basin of the cooling tower.
Cooling system auxiliary power consumption is site-specific and depends on the particulars of each design,
but as a very rough rule of thumb, the CW pumps and CT fans consume comparable amounts of power, and
their combination consumes about 1½ -2 % of the steam turbine’s output. Wet cooling towers consume
water at the rate of roughly 2-3 tons/MWh of steam turbine output.
15.1.6 COOLING TOWER PHYSICAL CONSTRUCTION
Figure 5. Counterflow (left) and Crossflow (right) cooling towers (courtesy of Marley Co.)
15-7
 M. A. Elmasri, 1990-2008
Chapter 15: Condensers & Cooling Towers
Two basic geometries are commonly used for cooling towers, illustrated in Figure 5. In the counterflow
arrangement, the water falls through a countercurrent of air, drawn upwards by the fan, and in the crossflow
arrangement the water falls through a crosscurrent of air, drawn horizontally into the tower across the falling
water, then upwards by the fan.
Fig. 6a. Splash-fill, suitable for dirty water,
but requires large tower volume
Fig. 6b. PVC Film-fill for clean water
Fig. 6c. Clog-resistant PVC Film-fill
Figure 6. Typical fill used in counterflow cooling towers (courtesy of Marley Co.)
Figure 7. Film fill with integral louvers for crossflow towers (courtesy of Marley Co.)
To maximise the area of contact between the air and water, the water is sprayed through a “fill”. The
simplest form of fill is a series of horizontal bars upon which the water impinges as it falls, slowing it down
and breaking it up. This is known as “splash fill”, and is shown in Figure 6a. Splash fill is simple and clogfree, but requires a fairly large volume within the cooling tower. To reduce tower volume, “film fill” is
used. This consists of corrugated PVC sheets with the water falling as a film flowing down the sheets,
slowed down by the tortuous corrugations, and the air being drawn through the narrow gaps between the
sheets. This creates a large contact area between air and water in a fairly compact volume, as shown in Fig.
6b. A more open form of film fill is used when water is contaminated, to minimise clogging by dirt and salt
deposits, as shown in Fig. 6c. A typical PVC film-fill geometry for crossflow towers is shown in Fig. 7.
15-8
 M. A. Elmasri, 1990-2008
Chapter 15: Condensers & Cooling Towers
Counterflow towers need an inlet plenum below the fill and an outlet plenum between the fill and the fan.
They tend to be higher, but with a more compact footprint, than crossflow towers. With their height, they
require more pump work to elevate the water, and with deep layers of fill, the fan power can also be higher
than for a crossflow design. The main argument in favour of counterflow is compactness of footprint, which
is especially important in large plants where the tower footprint is a major factor in site layout. The main
argument in favour of crossflow towers (with double flow, i.e. air entry from two sides) is the ability to
provide a large face area for the airflow, hence a large contact area without excessive fill depth or tower
height, reducing both fan and pump power.
The main structure of a cooling tower had, for many years, been commonly built of wood, mostly rotresistant species such as redwood. Other strong structural timbers such as Douglas fir are used. The wood is
usually pressure-treated with preservative to prevent fungi and bacteria from flourishing in the warm, humid
environment. While wood continues to be used, the trend is towards increasing use of modern synthetic
materials, such as FRP (Fiberglass Reinforced Plastic), which is more expensive but also more durable than
wood. Concrete or steel are also used as a durable structural materials, but they cost more than wood.
In power plants it is typical to install multiple cells, usually at least four, to allow some turndown by shutting
down some of the cells at low loads or cold ambients. If the plant is large, the number of cells would have to
exceed four anyway, since there is a maximum practical size for each cell. The largest common fan size is
roughly 10 meters (33 ft) in diameter, which corresponds to a maximum airflow of around 750 kg/s (1750
lb/s). With an L/G of 1.5, and an L/S (Condenser Steam) ratio of 45, this corresponds to a steam condensing
capacity of around 25 kg/s (55 lb/s) per cell. This corresponds to roughly 30 MW of steam turbine output.
Thus, one may expect at least one cooling tower cell per 30 MW of steam turbine output, based on fan
diameters limited to 10 meters (33 ft). Although practical fans have been made with diameters exceeding 15
meters (50 ft), which would allow up to 75 MW of steam turbine output per cooling tower cell, fan
diameters larger than 10 meters (33 ft) are uncommon.
Wood (Douglas Fir or Redwood, with
preservative)
Pultruded Fiberglass Composite (also called
FRP - Fiberglass Reinforced Plastic)
Figure 8. Typical cooling tower structures (courtesy of Marley Co.)
The least expensive fans used to draw air through the cooling tower would have fixed blade pitch and one
speed. With these fans, the only way to vary load efficiently is to turn off entire cells. For greater efficiency
at part-load, two-speed or variable-speed fans can be used, or alternatively, fixed speed fans with variable
15-9
 M. A. Elmasri, 1990-2008
Chapter 15: Condensers & Cooling Towers
blade pitch. Any of these solutions involves additional cost, but this should be evaluated against the reduced
power consumption at part-loads and cooler ambients enabled by the greater operating flexibility of the fans.
Fans are usually limited to a tip speed of about 70 m/s (230 ft/s). Since noise increases with tip speed, lownoise applications are limited to much lower speeds. With a typical tip speed of 60 m/s, a 10 m diameter fan
needs to rotate at about 115 RPM, and rotational speed varies in inverse proportion to diameter. Thus, fans
are driven through reduction gearing, since practical motors rotate at much higher speeds. Figure 9 shows a
typical cooling tower fan within its exit diffuser funnel, and a typical drive system, with a right angle
reduction gearing.
Figure 9. Typical cooling tower fans and drive system with reduction/right angle gearbox (courtesy of Marley Co.)
15.1.7 CONDENSATION PLUME
Figure 10. Condensation plume (courtesy of Marley Co.)
As the plume from a cooling tower mixes with the atmosphere, the mixing states fall along a line connecting
the initial plume state and the atmospheric state on the psychrometric chart. If the entire mixing line falls
within the superheated region of the chart (below the saturation line), no condensation will occur during
mixing and the plume will be invisible. If, on the other hand, portions of the line fall in the saturated zone
(above the saturation line), then condensation of vapour is likely to occur during the mixing process, and the
plume will be visible, as shown in Figure 10. Apart from the unsightly nature of the condensation plume,
causing displeasure to the populace, the plume can be hazardous if it causes wetness or icing on roads and
highways.
15-10
 M. A. Elmasri, 1990-2008
Chapter 15: Condensers & Cooling Towers
0.06
Humidity, lb/lb dry air, or kg/kg dry air
Saturation Line (100%
Relative Humidity)
0.05
90%
80%
A
70%
0.04
60% RH
50%
0.03
X
40%
0.02
30%
B
0.01
C
0
40
50
60
70
80
90
100
110
Dry-bulb Temperature (F)
Figure 11. Psychrometric chart, illustrating mixing between the tower discharge plume (A) and the atmosphere (B on a
warm day, and C on a cooler day)
As an example, if a wet cooling tower discharge is 100%-humid and at 100 ºF (38 ºC), point A on the
psychrometric chart of Figure 11, and the ambient is 50%-humid and at 80 ºF (27 ºC), point B on the chart,
no condensation will occur, and the plume would be invisible. If, on the other hand, the ambient were at
point C on the chart, 60 ºF (16 ºC) and 40% relative humidity, condensation will occur and the plume will be
partially visible. To help judge the likelihood of plume visibility, we use a Plume Visibility Index, defined as
the proportion of the mixing line which falls within the saturated region. This appears in the text output
section for the cooling tower in software programs, such as GT PRO. For mixing along the line A-C, the
plume visibility index is about 0.6, which is the ratio of the distance A-X to the distance A-C on the chart.
As may be seen from Fig. 11, plume condensation is most problematic on cool, humid days, when the
atmospheric state is as far to the left and as high up as possible on the psychrometric chart.
15.2
WET-DRY (“HYBRID”) COOLING TOWERS
One way to prevent plume condensation is to heat the 100%-humid plume as it leaves the tower. This may
be done by passing it over finned-tube coils placed above the tower, just under its fan. The warm water
entering the tower flows within these tubes before it is sprayed into the air. As seen in Figure 12, if the coils
heat the warm plume from A to A’ (by 7 °F or 4 °C in this example) at constant absolute humidity, the
relative humidity drops (to 80% in this example). When this new plume state mixes with the ambient air,
the entire mixing line A’-B is in the superheated region of the chart.
15-11
 M. A. Elmasri, 1990-2008
Chapter 15: Condensers & Cooling Towers
Humidity, lb/lb dry air, or kg/kg dry air
0.06
Saturation Line (100%
Relative Humidity)
0.05
A
90%
A’
80%
70%
0.04
60% RH
50%
0.03
40%
0.02
30%
0.01
B
0
40
50
60
70
80
90
100
110
Dry-bulb Temperature (F)
Figure 12. Psychrometric chart illustrating an example of plume abatement by a warm water coil in series with the wet
section, at its exit, heating the humid air before it leaves the cooling tower
Placing the warm water coils in series with the wet section, at its exit, where the temperature difference
between the water in the coils and the warm air leaving the wet section is minimal, requires a large coil
surface area, which introduces a large pressure drop in the air. An alternative way of abating the plume is to
place the warm water coils in parallel with the wet section, as illustrated by Figures 13 and 14. Ambient air
drawn through the dry section is heated up (B-B’), whilst cooling the warm water in the coils by a few
degrees. The warm water is then sprayed into the wet section. The warm humid air leaving the wet section
(A) is mixed with the heated, drier air (B’), to produce a mixture state A’, discharged from the tower. When
this plume mixes with the ambient air, the entire mixing line A’-B is in the superheated region of the chart.
Typically, the airflow through the dry section is 20%-50% of the total, but since it does not have the benefit
of the latent heat sink, the dry section heat transfer is only about 5-10% of the total heat rejection.
15-12
 M. A. Elmasri, 1990-2008
Chapter 15: Condensers & Cooling Towers
GT PRO 10.9 Maher Elmasri
Cooling System
Condenser
Exhaust steam
0.1665 M
82.31 T(WB)
79.49 %RH
5886 M
87.8 T
Wet-Dry Cooling Tower
A’
158.2 M
112 T
1.35 p
Hot CW
3867 M
106.9 T
Fans 577 kW
46.94 T(WB)
80 %RH
Air 1920.2 M
50 T
B’
A
3867 M
102.9 T
1920.2 M
83.28 T
B
3965 M
89.94 T
46.94 T(WB)
80 %RH
3870 M
50 T
B
3867 M
66.94 T
3867 M
66.94 T
Cold CW
CW Pump(s)
466.3 kW
158.3 M
112 T
1.35 p
119.6 M
Makeup
158.3 M
112.2 T
15.75 p
54.42 kW
Blowdown
23.91 M
to HRSG
Water Cooled Condenser and Wet-Dry Cooling Tower T-QDiagram
120
Exhaust Steam 1.35 psia
112
112
106.9
100
A’
A
TEMPERATURE [F]
Cooling Water
89.94
83.28
80
B’
66.94
Air
Air
60
46.94
50
B (wet bulb temperature)
Condenser 154680 BTU/s
CT 154680 BTU/s
40
0
20
40
60
B (dry bulb temperature)
80
100
120
140
160
180
200
HEAT TRANSFER [.001 X BTU/s]
Figure 13. Wet-Dry or “Plume Abatement” cooling tower, with dry section coil in parallel with the wet section
15-13
 M. A. Elmasri, 1990-2008
Chapter 15: Condensers & Cooling Towers
In the numerical example of Figures 13 and 14, the atmosphere (B) is at 50 °F and 80% humid. The wet
section discharge (A) is at 90 °F and 100% humid. Had this been a regular wet-only tower, most of the
mixing line A-B would have been in the saturated zone, and the total airflow to reject the heat would have
been 4300 lb/s. With the wet-dry design, the total required airflow is 5790 lb/s, i.e. about 35% greater.
About 1/3rd of the total airflow goes through the dry section, where it is heated to 83 °F at constant absolute
humidity, so it emerges at state B’ at about 25% relative humidity. Mixing this dry air with the wet section
discharge produces a tower plume at state A’, at 88 °F and about 80% relative humidity. When this plume
mixes with the atmosphere, along the line A’-B, the entire mixing process is in the superheated zone, so no
condensation plume occurs.
Humidity, lb/lb dry air, or kg/kg dry air
0.06
Saturation Line (100%
Relative Humidity)
0.05
80%
70%
0.04
60% RH
A
50%
0.03
A’
0.02
40%
30%
0.01
0
90%
B’
B
40
50
60
70
80
90
100
110
Dry-bulb Temperature (F)
Figure 14. Psychrometric chart illustrating process in Wet-Dry or “Plume Abatement” cooling tower, with dry section coil
in parallel with the wet section
In the numerical example illustrated above, the wet-dry tower consumes about 33% more fan power than an
equivalent wet-only tower. A wet-dry tower costs about twice as much as an equivalent, wet-only tower.
Recent designs of wet-dry towers use plastic fin-plate heat exchangers for the dry section, rather than
metallic finned-tube coils.
Figure 15 shows a wet-dry cooling tower, with louvers that can modulate, or shut-off, the airflow through
the dry section installed on top of the wet section. Reducing or shutting-off dry-section airflow on warm or
dry days, when there is no risk of plume condensation, saves fan power. Two-speed or variable-speed fans
should be used in conjunction with adjustable louvers on wet-dry towers to optimise efficient operation.
15-14
 M. A. Elmasri, 1990-2008
Chapter 15: Condensers & Cooling Towers
Figure 15. Wet-Dry or “Plume Abatement” cooling tower. The dry section has adjustable louvers, to conserve fan power
when conditions do not require full dry-section airflow for plume abatement (courtesy of Marley Co.)
15.3
NATURAL DRAFT WET COOLING TOWERS
Figure 16. Natural draft wet cooling towers (courtesy of Marley Co.)
Natural draft cooling towers use the buoyancy of the warm, humid plume to drive the airflow, saving the
auxiliary power consumed by the fans. Since the density difference between the warm plume and the
ambient is small, the tower needs to be tall to generate sufficient draft. For example, the density of typical
cooling tower discharge air, 100%-humid at 35 °C (95 °F) is 1.22 kg/m3. The density of ISO ambient, 60%humid at 15 °C (59 °F) is 1.121 kg/m3. Thus, a typical density difference between tower discharge and
ambient is on the order of 0.1 kg/m3. The draft created by this density difference is
15-15
 M. A. Elmasri, 1990-2008
Chapter 15: Condensers & Cooling Towers
∆P = ∆ρ g h ................................ (3)
where ∆P is the draft in Pascals, ∆ρ is the density difference in kg/m3, g is the acceleration of gravity (9.81
m/s2) and h the column height in meters.
From the above equation, the modest density difference of 0.1 kg/m3 has to act over a vertical column about
100 meters (330 ft) tall to generate a pressure difference of 1 millibar. The typical cooling tower needs
about 1-1.5 millibars (0.4-0.6 “ H2O) of draft to drive the air through the fill. Thus, without fans, a natural
draft tower needs to have a height around 100-150 m (330-500 ft).
For such a tall structure to be economical, it has to have a fairly thin shell. To be strong and stable, it needs
to be wide, which means a large airflow (at least 10,000 kg/s or 22,000 lb/s), hence its applicability is
limited to large plants, with steam turbines above 500 MW, say, implying combined cycles above 1500
MW. This size of plant is also needed to justify the additional capital expense, in terms of the savings in
auxiliary power consumption. At that size, the cost of a natural draft tower is about triple the cost of an
equivalent mechanical draft tower.
Natural draft towers are usually hyperbolic in profile, to create a strong, stable, structural shape, with a wide
base. The surface, a “hyperboloid of revolution” is actually composed of offset, angled, straight lines,
making it possible to build a very strong, interlocking mesh of straight reinforcing bars within a thin
concrete shell. This shape also lends itself to creating a smooth aerodynamic profile with an exit diffuser
that reduces exit kinetic energy loss. Such towers have been made of wood, but most large modern ones are
made of reinforced concrete. Some recent designs are made of Aluminum.
A natural draft tower works best in a cool, humid climate, which produces an adequate density difference
between the warm plume and the cool ambient, hence their widespread use in countries such as the UK. In
hot, dry climates, the heat sink is mostly latent, not sensible, so the density difference between the plume
and the ambient is small, which necessitates a taller tower. In fact, the plume within a mechanical draft
tower could be denser than the atmosphere in a hot, very dry climate, since the plume can be cooler than
ambient. Under such conditions, a mechanical draft tower’s fans will still create an air draft, but a natural
draft tower cannot be designed with these conditions.
15.4
DRY COOLING TOWERS
A dry cooling tower uses a flow of ambient air to cool the water flowing within finned tubes. Because the
air-side heat transfer coefficients are relatively low, the required surface area is very high, making such
towers quite expensive.
Dry cooling towers have been built both with mechanical draft fans and with natural draft. The construction
of the natural draft dry tower is very similar to that of a natural draft wet tower, except that the finned tube
arrays take the place of the wet tower fill. The tubes are arranged at the base, either as vertical tubes around
the inlet perimeter of the base, or as A-frames within the cylindrical tower structure near its bottom.
Table 3 shows a comparison of the main parameters for a wet and a dry cooling tower, based on ISO design
conditions, for a typical, reheat combined cycle with a 100-MW-nominal steam turbine. The dry mechanical
draft cooling tower may cost about $15m, whereas a wet mechanical draft cooling tower would only cost
about $2 m. The dry tower’s fan power consumption and footprint are both about four times the wet
tower’s.
Because the dry tower lacks the latent heat sink, its airflow is 4-5 times as much as a wet tower’s. With such
a large airflow, natural draft makes sense for smaller plants, with steam turbines as small as 100 MW, say,
whereas with wet towers natural draft only makes sense if the steam turbine is on the order of 500 MW or
larger.
15-16
 M. A. Elmasri, 1990-2008
Chapter 15: Condensers & Cooling Towers
GT PRO 10.9 Maher Elmasri
Cooling System
Dry Cooling Tower
Condenser
0.1662 M
Exhaust steam
156.5 M
104.9 T
1.1 p
59.72 T(WB)
29.65 %RH
30544 M
79.65 T
Fans 2700.9 kW
51.5 T(WB)
60 %RH
30544 M
59 T
9542 M
100 T
9542 M
84 T
9542 M
84 T
Hot CW
Cold CW
CW Pump(s)
884.9 kW
156.7 M
104.9 T
1.1 p
156.7 M
105.2 T
15.75 p
54.71 kW
to HRSG
p[psia], T[F], M[lb/s], Steam Properties: Thermoflow - STQUIK
Water Cooled Condenser and Dry Cooling Tower T Q Diagram
110
Exhaust Steam 1.1 psia
104.9
104.9
100
Cooling Water
90
TEMPERATURE [F]
84
79.65
70
Air
59
Condenser 152672 BTU/s
CT 152672 BTU/s
50
0
20
40
60
80
100
120
140
160
180
200
HEAT TRANSFER [.001 X BTU/s]
Figure 17. Dry Mechanical Draft Cooling Tower thermal parameters, sized at ISO conditions for a reheat combined
cycle with a 100-MW-nominal steam turbine
15-17
 M. A. Elmasri, 1990-2008
Chapter 15: Condensers & Cooling Towers
Table 3. Comparison of key overall parameters for Wet and Dry Mechanical Draft Cooling Towers, for a
typical reheat combined cycle with a 100-MW-nominal steam turbine @ ISO design conditions.
Parameter
Wet mechanical draft tower
Dry mechanical draft tower
48 mb (0.7 psia)
76 mb (1.1 psia)
96 MW
93 MW
Cooling Tower Cost, $m
2
15
Cooling Tower Footprint
800 m2 (9,000 ft2)
3,500 m2 (37,000 ft2)
800 kW
2,700 kW
Cooling Water Flow Rate
3,200 kg/s (7,000 lb/s)
4,300 kg/s (9,500 lb/s)
Cooling Tower Airflow
3,200 kg/s (7,000 lb/s)
14,000 kg/s (31,000 lb/s)
Typical Condenser Design Pressure
Steam Turbine Output
Cooling Tower Fan Power Consumption
15.4.1 COMPARISON WITH DRY AIR-COOLED CONDENSERS
The surface area, cost, and fan power consumption of a mechanical draft dry cooling tower are roughly
comparable to the same parameters for an equivalent dry air-cooled condenser. For the same condenser
pressure, the dry air-cooled condenser has the advantage of a higher temperature difference between the
condensing steam within the tubes and the ambient air, which allows it a slightly smaller surface area. The
dry air-cooled condenser also has the advantage that it is a single piece of major equipment, whereas the dry
cooling tower still needs a water-cooled surface condenser and cooling water pumps and piping. The main
disadvantage of the dry air-cooled condenser in this comparison is the need to pipe the steam, at high
vacuum, for a considerable distance between the turbine exhaust and the condenser cells. Even a very small
pressure drop in these pipes is significant because of the very low absolute pressure, and the low density of
the exhaust steam requires very large pipes. The steam piping, along with the large condenser surface area,
result in a large volume of the total hardware under vacuum, requiring large vacuum pumps and increasing
the possibility of air leaking into the system.
On balance, the dry air-cooled condenser will generally have a slight edge over a water-cooled surface
condenser coupled to a dry cooling tower.
15.4.2 THE HELLER SYSTEM
Figure 18 illustrates this system. It reduces the temperature difference between the condensing steam and
the ambient, by coupling the dry cooling tower to a direct contact spray condenser, rather than to a watercooled surface condenser. This eliminates the CW approach to condensate, shown in Fig. 3, and saves the
cost of the surface condenser tubing, improving system economics. Cooling tower surface area, footprint,
cooling system cost and auxiliary power consumption can all be reduced by 5-10%.
The condensate is circulated through the dry cooling tower, at a rate 40-50 times the steam flow rate, to sub
cool it by about 10 °C (18 °F). The subcooled condensate is sprayed into the steam, condensing it by
mixing. To avoid cavitation, and to minimise air ingress, the condensate should be pressurised to above
atmospheric pressure before sending it to the dry cooling tower. Rather than waste its excess pressure in a
reducing valve before spraying it back into the condenser vacuum, some systems use a hydraulic turbine to
recover this overpressure and help drive the condensate pump.
15-18
 M. A. Elmasri, 1990-2008
Chapter 15: Condensers & Cooling Towers
GT PRO 10.9 Maher Elmasri
Cooling System
Direct Contact
Condenser
Exhaust steam
0.1662 M
Dry Cooling Tower
28.39 %RH
28707 M
80.97 T
156.5 M
104.9 T
1.1 p
Fans 2541.7 kW
60 %RH
28707 M
59 T
6940 M
82.95 T
6940 M
104.9 T
156.7 M
104.9 T
1.1 p
CW Pump(s)
477 kW
156.7 M
105.2 T
15.75 p
54.71 kW
to HRSG
p[psia], T[F], M[lb/s], Steam Properties: Thermoflow - STQUIK
Heller SystemTQDiagram
110
Exhaust Steam 1.1 psia
104.9
104.9
104.9
Cooling Water
90
TEMPERATURE [F]
82.95
80.97
70
Air
59
Condenser 152672 BTU/s
CT 152672 BTU/s
50
0
20
40
60
80
100
120
140
160
180
200
HEAT TRANSFER [.001 XBTU/s]
Figure 18. Heller System (direct contact spray condenser with dry cooling tower) thermal parameters, sized at ISO
conditions for a reheat combined cycle with a 100-MW-nominal steam turbine
15-19
 M. A. Elmasri, 1990-2008
Chapter 15: Condensers & Cooling Towers
15.5
DRY AIR-COOLED CONDENSER
Figure 19. Dry Air-cooled Condenser at a 600 MW Cogen CC (Courtesy of GEA Power Cooling Systems)
Wet cooling towers consume water at the rate of roughly 2-3 tons/MWh of steam turbine output. Thus, a
100 MW steam turbine using a wet cooling tower (or other evaporative cooling system like a wet-surface
condenser) will consume about 200-300 tons/hr of water, or about 2 million tons per year (2 million cu.m.).
In addition, the cooling water needs to be treated against algae and bacteria, and the chemicals used for the
purpose may be environmentally objectionable. Disposing of the blowdown water, with concentrated
impurities, is another source of environmental costs. Dry cooling systems do not consume any water and
avoid these problems, but are much more expensive, both in initial capital and in auxiliary power
consumption.
Figure 20. Dry Air-cooled Condenser being built for a 500 MW CC, showing “A” Frames and steam distribution headers
(Courtesy of GEA Power Cooling Systems)
15-20
 M. A. Elmasri, 1990-2008
Chapter 15: Condensers & Cooling Towers
In a dry air-cooled condenser, the steam is distributed through large headers to a series of cells. Each cell
has two banks of tubes, arranged in the shape of an inverted “V”, with the steam distribution header at the
top. Fans force a draft of air through the tube banks, from the open side of the “V” at the bottom. The steam
condenses inside these tube, and the condensate is gathered in the headers at their bottom. Because of the
very high condensing heat transfer coefficient on the inside, and the much smaller coefficient between air
and tubes on the outside, the outside surface is densely finned to increase its surface area. Typically, the
outside finned surface area may be about fifty times as large as the inside surface area. The tubes may be
round or oval in cross-section, and they may be made of galvanised steel or Aluminum. Proponents of
Aluminum have claimed that Zinc washes slowly off the galvanised steel finned tubes and results in a
poisonous solution draining into the ground.
Figure 21. Fan deck for an ACC, with regular fans (Courtesy of GEA Power Cooling Systems)
Figure 22. Ultra-low-noise ACC fan (Courtesy of GEA Power Cooling Systems)
15-21
 M. A. Elmasri, 1990-2008
Chapter 15: Condensers & Cooling Towers
Fans, whether in an air-cooled condenser or in a cooling tower are the principal source of noise. In cooling
towers, the falling water produces additional noise as it splashes around. Fan noise can be reduced by using
lower tip speeds on the fans, but this implies a lower pressure rise. Low pressure rise means a thinner tube
bundle on an ACC or thinner fill on a CT, which, in turn, means more frontal area, hence more cells and
higher capital cost. Advances in fan aerodynamic design allows greater pressure rise with lower speeds and
lower noise, using blade shapes such as shown in Fig. 22.
Table 1 shows some key parameters for water-cooled and air-cooled condensers, as well as for heat recovery
boilers. Due to the very low heat transfer coefficient on its air-side, the overall heat transfer coefficient in an
ACC is comparable with that for a typical heat recovery boiler. Total required surface area exceeds that of
the boiler, due to the lower mean temperature difference employed in the air-cooled condenser. Total cost of
the ACC and the HRSG are comparable.
As shown in Table 1, an ACC costs about ten times as much as a water-cooled surface condenser. However,
the ACC is stand-alone, whereas the WCC still needs CW piping, pumps and a cooling tower, usually of the
wet mechanical draft type. When one considers these costs, one finds that the typical dry ACC costs about
three to four times as much as the typical wet cooling system. In addition, using an ACC would typically
necessitate a higher condenser pressure, reducing ST output, and the ACC would consume more auxiliary
power than the wet cooling system. As a consequence, using an ACC instead of a wet cooling system in a
combined cycle would result in a total plant net power output reduction of 1¼ - 2½ %, as well as a total
plant cost increase of 5-10%. This is the cost of virtually eliminating the combined cycle’s water
consumption.
15.6
WET-SURFACE CONDENSER
This relatively uncommon system is like an air-cooled condenser in that the steam condenses inside cooled
tubes, but is like a wet-cooling tower in that the heat rejection is largely by evaporation. The condenser
tubes are horizontal and unfinned, usually made of stainless steel of about 2” (50.8 mm) diameter. Rather
than fins to enhance heat transfer to the air, water is sprayed onto the tubes to soak their surfaces, and air is
drawn over the wet surface by fans.
Figure 23a. Wet-surface condenser (photo by the author courtesy of Masspower)
15-22
 M. A. Elmasri, 1990-2008
Chapter 15: Condensers & Cooling Towers
Figures 23b & 23c. Wet-surface condenser (photo by the author courtesy of Masspower)
The water and auxiliary power consumption of a wet-surface condenser are roughly the same as for a watercooled condenser with its wet cooling tower and associated CW pumps and pipes. The wet-surface
condenser may cost about 30% more, however, but is still a fraction of the cost of a dry air-cooled
condenser. The chief advantage of the wet surface condenser is that some of its cells may be run dry,
without any water spray, and will still contribute to the condensing load by air-cooling, especially on cold
days. Thus, it can use partial air-cooling during times of water shortage, albeit at the expense of higher
condensing pressure and less ST power. A conventional water-cooled condenser with a cooling tower has
no ability to reject any portion of the heat without consuming water.
15-23
 M. A. Elmasri, 1990-2008
Chapter 15: Condensers & Cooling Towers
15.7
HYBRID COOLING SYSTEMS
In areas of potential water shortage, both an air-cooled condenser and a water-cooled condenser/cooling
tower may be used in parallel. If, for instance, each is designed to carry half the load, the water
consumption at design conditions will be halved, relative to a system with only wet cooling. Further, at
times of water shortage, or in very cold ambients, some of the wet tower cells can be shut down, raising
condenser pressure and transferring more load to the ACC. One manufacturer, GEA, designs and supplies
such a system under the trade name of a GEA “pac system”.
Figure 24. Hybrid cooling system with dry and wet cooling in parallel
15-24
 M. A. Elmasri, 1990-2002
Chapter 16: CC Off-Design Behaviour
OFF-DESIGN BEHAVIOUR OF COMBINED CYCLES
Revised November, 2002
© Maher Elmasri 1990-2002
16.1
OFF-DESIGN COMPONENT INTERACTIONS IN A COMBINED CYCLE
In a combined cycle, the thermodynamic boundary conditions on each component are imposed by the
other components with which it interacts. These boundary conditions, such as pressures and
temperatures, depend on the mass flow rates passing from one component to the next. Since these mass
flow rates themselves are dependent on the boundary conditions, an iterative solution is needed to
simultaneously solve for the plant mass flow rates and thermodynamic conditions entering and leaving
each component. The interactions between the principal components are summarised below.
16.1.1 INTERACTIONS BETWEEN GAS TURBINE & HEAT RECOVERY BOILER
The hot gas entering the boiler determines its driving force, dictating the mass flow it can generate as a
function of the pressure it sees at its exit. For a given hot gas mass flow rate, composition, and
temperature, the HRB has a pumping characteristic as shown in Figure 3 of Chapter 11.
There is some minor feedback from the boiler to the gas turbine as well. At off-design, the HRB draft
loss varies as a function of exhaust gas mass flow rate, composition, initial temperature, supplementary
firing level (if any), and rate of heat absorption in the various sections of the boiler. The varying draft
loss influences gas turbine performance. A higher draft loss leads to less gas turbine power output, as
well as hotter gas turbine exhaust, which, in turn, affects boiler behaviour.
16.1.2 INTERACTIONS BETWEEN HEAT RECOVERY BOILER & STEAM TURBINE
In a combined cycle, the steam turbine imposes the flow resistance upon the HRB exit port(s). Thus, on a
pressure-mass flow diagram, the intersection of the steam turbine resistance characteristic with the boiler
pumping characteristic determines the operating condition. This is illustrated in Figure 1. With the boiler
driven by flue gases at the flow rate and temperature corresponding to full gas turbine load, it has
pumping characteristic (a) on Figure 1. The steam turbine at full-load and VWO (Valves Wide Open)
exhibits a flow resistance characterized by a linear relationship between pressure and massflow, as
described above, indicated by the line a-b-c-d-e on Figure 1. This line is sometimes called the "Willans
Line". The intersection of the two curves sets the operating conditions, just as with any pump/pipe
system.
If the energy of the flue gas entering the boiler is reduced, by reducing the combination of exhaust gas
mass flow rate and temperature, such as by reducing gas turbine load, or by turning a duct burner down,
the boiler pumping characteristic shifts down and to the left of Figure 1, as discussed in Chapter 11. If the
steam turbine were operated as a passive "slave", with valves left wide open, its resistance characteristic
is unaltered. The result would be for the boiler/steam turbine combination to operate along the Willans
Line, a-b-c-d-e, in "Sliding Pressure" mode, as the flue gas energy is reduced.
Suppose that the pressures below point (c) on Fig. 1 were too low for the boiler to operate satisfactorily,
due to excessive steam velocities and pressure drops, as well as due to economiser steaming. Then the
steam turbine resistance would need to be increased to hold up the boiler pressure. This could be effected
by adjusting the opening of a throttle valve feeding the steam turbine, so that the operating points d' and e'
would be reached rather than d and e.
16-1
 M. A. Elmasri, 1990-2002
Chapter 16: CC Off-Design Behaviour
It is fairly common to operate combined cycles in sliding pressure, down to about 75% of nominal, fullload pressure, then to hold up the pressure by shutting the steam turbine valve(s) as shown on Fig. 1.
te
rac
e'
Closing throttle shifts
Resistance Characteristic
in this direction
a
d'
b
c
d
e
tic
i
t er
rac
ha tion
g C ndi
pin co
um inal
B P om
HR at n
Steam Pressure
a
Ch
nce O
a
t
sis VW
Re at
ST
ri s
sti
c
Reducing exhaust energy shifts
HRB Pumping Characteristic in this direction
Steam Mass Flow Rate
Figure 1. Combined cycle operating point is determined by the intersection of the boiler pumping curve with the
steam turbine resistance curve
16.1.3 INTERACTIONS BETWEEN STEAM TURBINE & CONDENSER
Just as with the HRB/ST interface, the steam turbine and condenser interact to determine the operating
condition of both. For a given steam flow rate, changing the condenser vacuum by varying cooling water
flow rate or temperature affects steam turbine performance and exhaust steam enthalpy. The change in
exhaust steam enthalpy, in turn, influences the heat rejection rate, and hence the condenser vacuum.
Low condenser pressure lowers steam density and increases steam turbine exhaust annulus velocity and
vice versa. The variation in annulus velocity affects steam turbine exhaust loss, as discussed in §14.4.4 &
§14.4.5. It should be borne in mind that exhaust loss typically represents 1-6% of the expansion work in
a condensing steam turbine. Thus, at very cold ambients, reducing condenser pressure to the minimum
attainable value may increase the exhaust loss to the point that it nullifies, or even exceeds, the benefit of
a lower condenser pressure. This is particularly true with air-cooled condensers, due to their relatively
high parasitic loss to drive the fans.
The speed of sound in steam at typical steam turbine exhaust temperatures is about 410 m/s (1345 ft/s).
The useful limit to decreasing steam turbine exhaust pressure with a cold ambient is the point where sonic
velocity is reached in the steam turbine exhaust annulus. Further reduction in condenser pressure creates
a shock wave in the exhaust annulus. This prevents the lower condenser pressure from propagating
upstream to the exit plane of the last turbine wheel, which does not see the lower condenser pressure.
16.1.4 INTERACTIONS BETWEEN CONDENSER & HRB
Changes in condenser pressure will affect the feedwater temperature supplied to the boiler’s coolest heat
exchanger, the low temperature economiser (or condensate preheater). This affects the stack temperature,
the water temperature reaching the deaerator, and thus the integral deaerator pressure if it is free to float.
16-2
 M. A. Elmasri, 1990-2002
Chapter 16: CC Off-Design Behaviour
If the deaerator is heated or pegged using steam from another source, the demand for pegging steam will
increase with a lower condenser pressure and temperature.
16.2
EFFECT
LOAD
OF
AMBIENT TEMPERATURE
ON
COMBINED CYCLES
AT
FULL
16.2.1 DISCUSSION OF GENERAL TRENDS
The effect of ambient temperature on gas turbine performance has been discussed in detail in §6.3. For
most gas turbines, a cooler ambient results in higher power output, better gas turbine heat rate, higher
exhaust gas mass flow rate, and lower exhaust gas temperature. Table 1 in §6.3.2 provides approximate
numerical guidelines for these effects, applicable to most heavy duty gas turbines.
Ambient temperature affects the steam bottoming cycle by changing conditions at both its heat source, the
GT exhaust, and its heat sink, the environment. In the absence of supplementary firing, the changing gas
turbine exhaust conditions affect boiler steam production rates, and possibly affect steam temperatures as
well. The ambient environment affects condenser pressure in a way that depends on the type of cooling
system.
Whereas a cold ambient can significantly improve gas turbine performance, its beneficial effect on the
steam bottoming cycle is much weaker. First, the bottoming cycle’s hot side is adversely affected, since
its heat recovery efficiency falls with the larger flow rate of cooler exhaust gas, and its steam
temperatures may fall as well. Second, the bottoming cycle’s cool side cannot take full advantage of a
cold ambient, since cooling water cannot drop below freezing, and since excessively low condenser
pressures lead to high steam turbine exhaust annulus velocities, with correspondingly high exhaust losses.
This latter point bears additional discussion. Choice of steam as the working fluid in a bottoming cycle
precludes the benefits of cold ambients, below freezing temperatures. As the condenser temperature
approaches the triple-point, at which steam, water, and ice coexist (close to 0 °C or 32 °F), the saturation
pressure approaches zero and steam specific volume approaches infinity1. Whereas the additional gain in
steam enthalpy drop is approximately linear with reduction in condenser saturation temperature, the
increase in steam specific volume is exponential. Thus, the increase in steam velocity in an annulus of
given area is exponential, and exit kinetic energy is proportional to the square of velocity. To numerically
illustrate this fact, consider a steam turbine annulus sized for an exhaust velocity of 200 m/s (656 ft/s) at a
pressure of 60 mb (0.87 psia). A drop in condenser pressure by a factor of two, to 30 mb (0.44 psia),
corresponds to a reduction in saturation temperature from 36 °C (97 °F) to 24 °C (75 °F), with an
additional expansion line enthalpy drop of about 85 kJ/kg (37 BTU/lb). However, annulus velocity
doubles, and exhaust kinetic energy quadruples, from 20 kJ/kg (8.6 BTU/lb) to 80 kJ/kg (34.4 BTU/lb).
Thus, 70% of the gain in expansion line enthalpy drop is lost in additional exhaust kinetic energy.
Further reductions in condenser pressure would be counterproductive, and the only way to take full
advantage of sub-freezing ambients is to use a bottoming cycle working fluid other than water.
A hot ambient degrades gas turbine performance but is beneficial to the bottoming cycle’s hot side, since
its heat recovery efficiency improves with the lower flow rate of hotter exhaust gas. The hot ambient,
however, increases condenser pressure, so is detrimental to the bottoming cycle’s cool side.
16.2.2 EFFECT ON TYPICAL UNFIRED COMBINED CYCLE WITHOUT GT INLET COOLING
The discussions in §16.2.1 are illustrated through a numerical example of a 250-MW-class triple-pressure
reheat combined cycle, based on a modern heavy duty gas turbine. The plant is designed and optimised
In a practical, not a numerical sense, since vapour pressure over ice is very small, but still finite, and although the
corresponding specific volume is very large, it is not infinite
1
16-3
 M. A. Elmasri, 1990-2002
Chapter 16: CC Off-Design Behaviour
for ISO conditions, and its correction curves at full-load are shown as a function of ambient, at a constant
relative humidity of 60%. No supplementary firing and no GT inlet air cooling is assumed, but the effects
of evaporative inlet air cooling are shown in §16.2.3 below. Although the results shown are
representative of typical plants of this sort, precise numerical values will vary from one plant to another,
depending on the specific design assumptions, equipment sizing philosophy, and operating procedures.
When comparing the correction curves shown in this section for plants with different cooling systems, it
is important to remember that the correction curves are not absolute results. They are simply factors to be
applied to the absolute performance of each plant type at its design point, taken to be ISO conditions in
this section. Thus, if plant ‘A’ exhibits a higher output correction factor at a hot ambient than plant ‘B’,
this does not necessarily mean that plant ‘A’ will produce more power on a hot day, since plant ‘B’ may
produce so much more power at ISO conditions, that even with its smaller hot-day correction factor it
produces more power than plant ‘A’ on a hot day.
16.2.2.1 Combined Cycle with Water Cooled Condenser & Mechanical Draft Cooling Tower
Figures 2 and 3 show how power and heat rate are affected by ambient temperature, and Figure 4 helps to
understand these trends. The gas turbine correction curve is substantially similar to those discussed in
§6.3, and the only effects from the equipment around the gas turbine is the variation in inlet and exhaust
pressure drops with flow conditions. At cooler ambients, the higher mass flow rate of air and exhaust
gases result in higher pressure drops, and vise versa. This slightly flattens the gas turbine correction
curves compared to those based on given, constant, inlet and exhaust pressure losses.
As discussed in §16.2.1, the gain in steam turbine output is modest at cold ambients, mostly due to the
declining heat recovery boiler efficiency. Figure 4, shows that HRB efficiency drops by over four
percentage points, from 86.6% at the ISO design point to 82.2% at the –15 °C (5 °F) ambient. This is
caused by the drop in GT exhaust temperature, by about 18 °C (32 °F) in this case, and the increase in
exhaust mass flow rate, by about 10% in this case. Each of these effects, on its own, is detrimental to
HRB efficiency, as explained in §11.1.2 and §11.1.3, respectively.
Effect of Ambient on Typical 3PRH CC - WCC/MDCT
5
23
41
59 F
77
95
113
Power Output, % of Design Point
115
GT Power
110
ST Power
CC Net Output
105
100
95
90
85
80
75
-15
-5
5
15
25
35
45
Ambient Temperature, C
Figure 2. Power correction curves for a typical triple-pressure reheat
combined cycle with mechanical draft cooling tower
16-4
 M. A. Elmasri, 1990-2002
Chapter 16: CC Off-Design Behaviour
Effect of Ambient on Typical 3PRH CC - WCC/MDCT
112
5
23
41
59 F
77
95
113
GT Heat Rate
Heat Rate, % of Design Point
110
ST Pseudo Heat Rate
CC Net Heat rate
108
106
104
102
100
98
96
-15
-5
5
15
25
35
45
Ambient Temperature, C
Figure 3. Heat rate correction curves for a typical triple-pressure reheat
combined cycle with mechanical draft cooling tower
Figure 4 also shows that the steam bottoming cycle internal efficiency declines at colder ambients. This
is primarily due to the slide in steam temperatures at the throttle and at the hot reheat. This slide, of about
16 °C (29 °F) could have been mitigated by oversizing the superheater and reheater at the ISO design
point, but this would have led to greater desuperheating flows, with consequent loss of efficiency, at
warmer ambients. Additionally, the steam turbine exhaust increases significantly (more than triples in
this example) between the ISO design point and the coldest ambient. Again, this could have been
mitigated by oversizing the steam turbine exhaust, but in addition to the extra cost, this would reduce
performance on warm days. The combination of lower boiler efficiency and lower steam cycle internal
efficiency at cold ambients leads to a lower overall bottoming cycle efficiency, as seen in Figure 4.
Effect of Ambient on Typical 3PRH Bottoming Cycle - WCC/MDCT
5
23
41
59 F
77
95
113
38
92
37
90
35
34
88
33
32
86
31
30
29
HRB efficiency, %
Steam cycle efficiencies, %
36
84
Steam cycle internal efficiency
HRSG efficiency
28
Steam cycle overall efficiency
82
27
26
-15
-5
5
15
25
35
45
Ambient Temperature, C
Figure 4. Effect of ambient on main bottoming cycle efficiency parameters for a typical 250-MW-class combined
cycle with a water cooled condenser and mechanical draft cooling tower
16-5
 M. A. Elmasri, 1990-2002
Chapter 16: CC Off-Design Behaviour
Despite the drop in overall bottoming cycle efficiency, the steam turbine output still increases at cold
ambients, as shown in Figure 2, since the total energy input to the steam cycle is still higher on a cold
day, due to the increased exhaust gas flow rate, and in spite of the declining exhaust gas temperature. The
steam turbine pseudo heat rate, however, worsens significantly, as shown in Figure 3, since the meager
increase in steam turbine output is disproportionate to the significant increase in gas turbine fuel input. At
cold ambients, the worsening steam turbine pseudo heat rate overwhelms the improvement in gas turbine
heat rate, with the net result of a worsening combined cycle heat rate, as seen in Figure 3.
The trends described above for the boiler efficiency are essentially reversed on hot days, where the
increase in GT exhaust temperature and reduction in mass flow rate both work to improve HRB
efficiency. The improved boiler efficiency mitigates the decline in steam cycle internal efficiency
occasioned by the higher condenser pressure. Steam turbine output declines in unison with gas turbine
output, as shown in Figure 2, but the worsening steam turbine pseudo heat rate is not quite as severe as
the degradation of gas turbine heat rate. Hence, net plant heat rate correction is less severe than the gas
turbine’s, as seen in Figure 3.
16.2.2.2 Combined Cycle with Open-Loop Water Cooled Condenser
Generally, open-loop cooling water sources have a narrower range of temperature fluctuation than the
ambient air, or than water cooled by ambient air in a wet cooling tower. Thus, a combined cycle cooled
in this fashion would have flatter correction curves. These are illustrated in Figures 5 and 6 for the same
combined cycle discussed in §16.2.2.1, but designed at ISO conditions with an open-loop condenser
cooled by seawater. It is assumed that the cooling water temperature changes by 4° for each 10° change
in ambient air temperature.
Effect of Ambient on Typical 3PRH CC - Open Loop WCC
5
23
41
59 F
77
95
113
Power Output, % of Design Point
115
GT Power
110
ST Power
CC Net Output
105
100
95
90
85
80
75
-15
-5
5
15
25
35
45
Ambient Temperature, C
Figure 5. Power correction curves for typical triple-pressure reheat
combined cycle with open-loop water cooling
For cold ambients, the correction curves are generally similar to those obtained with a wet cooling tower.
For hot ambients, however, performance is much better than with a wet cooling tower, since the cooling
water temperature does not rise as much as the cooling tower’s. This prevents the worsening steam
turbine pseudo heat rate seen in Figure 3, as well as the precipitous decline in steam cycle internal
efficiency seen in Figure 4 for the wet cooling tower.
16-6
 M. A. Elmasri, 1990-2002
Chapter 16: CC Off-Design Behaviour
Effect of Ambient on Typical 3PRH CC - Open Loop WCC
5
23
41
59 F
77
95
113
112
GT Heat Rate
Heat Rate, % of Design Point
110
ST Pseudo Heat Rate
CC Net Heat rate
108
106
104
102
100
98
96
-15
-5
5
15
25
35
45
Ambient Temperature, C
Figure 6. Heat rate correction curves for typical triple-pressure reheat
combined cycle with open-loop water cooling
Figure 7 shows that for the open loop cooling system, the improvement in HRB efficiency virtually
cancels the decline in steam cycle internal efficiency at hot ambients, leading to a nearly flat bottoming
cycle overall efficiency at hot ambients, when compared with the wet cooling tower curve of Figure 4.
Effect of Ambient on Typical 3PRH Bottoming Cycle - Open Loop WCC
5
23
41
59 F
77
95
113
38
92
37
90
35
34
33
88
32
86
31
30
29
HRB efficiency, %
Steam cycle efficiencies, %
36
84
Steam cycle internal efficiency
HRSG efficiency
28
Steam cycle overall efficiency
27
82
26
-15
-5
5
15
25
35
45
Ambient Temperature, C
Figure 7. Effect of ambient on main bottoming cycle efficiency parameters for a typical 250-MW-class combined
cycle with an open-loop water cooled condenser
16.2.2.3 Combined Cycle with Dry Air Cooled Condenser
Figures 8 and 9 show the correction curves, and Figure 10 shows the key bottoming cycle efficiencies.
Cold day correction factors are quite similar to the wet cooling tower and the open loop system, with the
air-cooled condenser showing a slightly greater cold-day performance enhancement relative to its ISO
performance.
16-7
 M. A. Elmasri, 1990-2002
Chapter 16: CC Off-Design Behaviour
Hot day performance, however, is much worse with the air-cooled condenser, due to the significant rise in
condenser pressure. When comparing the figures for the air-cooled condenser with their equivalents for
water-cooled plants, please note the different y-axis scale used for the air-cooled cases, due to their poor
hot-day performance.
Effect of Ambient on Typical 3PRH CC - ACC
5
23
41
59 F
77
95
113
115
GT Power
Power Output, % of Design Point
110
ST Power
CC Net Output
105
100
95
90
85
80
75
70
-15
-5
5
25
15
35
45
Ambient Temperature, C
Figure 8. Power correction curves for typical triple-pressure reheat
combined cycle with dry, air-cooled condenser
Effect of Ambient on Typical 3PRH CC - ACC
5
23
41
59 F
77
95
113
15
25
35
45
Heat Rate, % of Design Point
118
116
GT Heat Rate
114
CC Net Heat rate
ST Pseudo Heat Rate
112
110
108
106
104
102
100
98
96
-15
-5
5
Ambient Temperature, C
Figure 9. Heat rate correction curves for typical triple-pressure reheat
combined cycle with dry, air-cooled condenser
16-8
 M. A. Elmasri, 1990-2002
Chapter 16: CC Off-Design Behaviour
Effect of Ambient on Typical 3PRH Bottoming Cycle - ACC
5
23
41
59 F
77
95
113
38
Steam cycle internal efficiency
37
Steam cycle overall efficiency
90
35
HRB efficiency, %
Steam cycle efficiencies, %
92
HRSG efficiency
36
34
88
33
32
86
31
30
84
29
28
82
27
26
-15
-5
5
15
25
35
45
Ambient Temperature, C
Figure 10. Effect of ambient on main bottoming cycle efficiency parameters for a typical 250-MW-class combined
cycle with a dry, air-cooled condenser
16.2.3 UNFIRED COMBINED CYCLE WITH GT EVAPORATIVE INLET AIR COOLING
Figure 11 illustrates the effects of gas turbine evaporative inlet air cooling on improving the correction
curves at hot ambient conditions. This is for the plant with a water cooled condenser and wet cooling
tower, identical in all respects to the model whose results are shown in Figures 2 through 4, except for the
addition of the evaporative inlet cooler. This is assumed to be 95% effective, and to be turned on only at
ambient temperatures above ISO.
Effect of Ambient on Typical 3PRH CC - WCC/MDCT - Evap Inlet Cooler
Power & Heat Rate, % of Design Point
115
5
23
41
59 F
77
95
113
5
15
25
35
45
110
105
100
95
90
GT Power
85
ST Power
CC Net Output
80
CC Net Heat rate
75
-15
-5
Ambient Temperature, C
Figure 11. Power and net combined cycle heat rate correction curves for a typical 250-MW-class combined cycle
with GT inlet air evaporative cooling, water-cooled condenser, and wet mechanical draft cooling tower
Comparing Figure 11 with Fig. 2 shows that the evaporative cooler allows the plant to produce 85% of its
net ISO output at the hottest ambient (45 °C, 113 °F), whereas in its absence, it could only produce 80%.
16-9
 M. A. Elmasri, 1990-2002
Chapter 16: CC Off-Design Behaviour
Comparing Figure 11 with Fig. 3 shows that the evaporative cooler improves plant heat rate at the hottest
ambient (45 °C, 113 °F), but only very slightly, from 107.3% of ISO heat rate, to 106% of ISO heat rate.
This is accordance with the discussions given in §13.2.1.3.
16.2.4 SUPPLEMENTARY FIRING TO MAINTAIN OUTPUT AT HOT AMBIENTS
Many plants are designed with supplementary firing to increase output at hot ambients. This is effective,
but costly, both in equipment first cost and in heat rate, because supplementary firing just raises plant
output through the steam turbine, which is the weaker contributor. Measures that raise output of the gas
turbine, the stronger contributor, are generally more cost effective.
To estimate the extent of supplementary firing needed to preserve output on a hot day, we first note that in
a typical combined cycle, the gas turbine produces twice the power of the steam turbine, so if the gas
turbine produces 100 units of output, the steam turbine produces 50, and the plant produces 150. As
shown in §6.3.2, a typical heavy duty gas turbine loses output at the rate of roughly 7% per 10 °C (18 °F)
increase in inlet air temperature, so its output drops from 100 to 93, which means that the steam turbine
output must be increased from 50 to 57 to preserve the plant’s total output. Thus, the steam turbine’s
power needs to be increased by about 14% per 10 °C increase in ambient. This requires an increase in
boiler steam production by more than 14%, say 15%, since condenser pressure will rise with the hotter
ambient and the increased steam condensation load imposed upon the cooling system. The gas turbine
exhaust mass flow rate falls roughly 4% per 10 °C increase in ambient, so the boiler needs to generate
15% more steam from 4% less exhaust gas, i.e. the steam generation rate per unit of exhaust gas flow rate
needs to increase by 19%. Recalling the concept of the heat recovery boiler’s Effective Temperature
Difference, defined in §9.3, and noting that for a modern F-class gas turbine it is about 300 °C (540 °F),
between the gas and the HP saturation at unfired, ISO conditions, we can conclude that we need to
increase the gas temperature by about 0.19*300=57 °C to increase the HP flow by 15%. In a triplepressure boiler, however, increasing the HP steam generation by supplementary firing diminishes the IP
and LP steam production rates, so the increase in HP needs to be even greater, to compensate for the fall
in IP and LP, say 18%, not 15%. Thus, we can conclude that in a triple-pressure reheat cycle, we need to
raise the effective temperature difference by about 22%, or about 0.22*300=66 °C to maintain ISO output
if the ambient were to increase by 10 °C above ISO. Since the gas turbine’s exhaust temperature rises by
only about 7 °C per 10 °C increase in ambient, roughly 60 °C of temperature rise needs to be created by
the duct burner.
As the ambient gets higher, the need for supplementary firing increases at a rate greater than linear, since
the percentage increase needed in effective temperature difference is applied to a larger, base value. Also,
if the condenser is cooled by evaporative cooling towers, cooling water temperature increases at a rate
greater than linear with ambient dry bulb temperature, if relative humidity is constant or increasing.
16-10
 M. A. Elmasri, 1990-2002
Chapter 16: CC Off-Design Behaviour
Typical 3PRH CC - WCC/MDCT - DB for flat output
5
23
41
59 F
77
95
113
5
15
25
35
45
Power & Heat Rate, % of Design Point
140
135
GT Power
ST Power
130
CC Net Output
125
CC Net Heat rate
120
115
110
105
100
95
90
85
80
75
-15
-5
Ambient Temperature, C
Figure 12. Power and net combined cycle heat rate correction curves for typical 250-MW-class combined cycle
supplementary firing for flat output at ambients above ISO (water-cooled, wet mechanical draft cooling tower)
Figure 12 shows the performance of the same triple-pressure reheat combined cycle used previously,
except that here it is supplementary fired to the level needed to maintain the output at hot ambients the
same as at ISO. It shows that to produce the nominal ISO output at 45 °C (113 °F) ambient, the steam
turbine must produce almost 40% more power than at ISO. If the burner is at the inlet to the boiler, it
would have an exit temperature of about 910 °C (1670 °F) in this example. This is hardly practical. First,
this level of supplementary firing is higher than common practice, and would require ceramic lining in the
boiler duct. This, on its own, is not very problematic, and the duct burner could be located within the
boiler, after the gas has been cooled by some of the superheaters and reheaters. The real difficulty is with
the steam turbine. To produce 40% more power, it needs a throttle flow more than 40% greater than at
ISO, because the IP and LP flow rates drop with supplementary firing, and because the condenser
pressure rises at hot ambients and with increasing steam flow. In this example, it turns out that it takes
almost 60% more steam at the throttle to get 40% more power at the hottest ambient. If the steam turbine
is optimised to run with valves wide open at ISO conditions, its inlet pressure will climb by 60%, so if its
ISO design point were selected at a reasonable pressure, in the 100 bar range, say, it would have to be
capable of 160 bars. This would make the plant much more expensive, because not just the steam turbine,
but the boiler, piping, and feedpumps as well, all have to be capable of the higher pressure. Alternatively,
the steam turbine may be designed with the capacity to swallow the throttle flow rate of the fully fired
condition, at its nominal pressure, but then its ISO efficiency would be compromised by the need for
partial closure of the control valves at ISO, to maintain boiler pressure. Since, presumably, the plant
operates many more hours at ambients near ISO than at very hot ambients, any sacrifice of efficiency at
ISO is also expensive on a year-round basis.
The conclusion is that whereas it is perfectly possible to maintain output on hot ambients by
supplementary firing alone, it is expensive, both in equipment first cost, and in compromising efficiency,
or both. A more suitable approach is to focus on enhancing gas turbine performance on hot days, since,
after all, it is the gas turbine which produces most of the plant’s output. This can be done by inlet air
evaporative cooling, chilling, or overspray fogging, all of which have been discussed in Chapter 13. If
needed, these measures can be combined with supplementary firing at more modest levels to achieve even
greater power outputs.
16-11
 M. A. Elmasri, 1990-2002
Chapter 16: CC Off-Design Behaviour
Typical 3PRH CC Duct Fired for Flat Output with Evap Cooler
5
23
41
59 F
77
95
113
5
15
25
35
45
Power & Heat rate, % of Design Point
130
GT Power
125
ST Power
CC Net Output
120
CC Net Heat rate
115
110
105
100
95
90
85
-15
-5
Ambient Temperature, C
Figure 13. Power and net combined cycle heat rate correction curves for typical 250-MW-class combined cycle with
GT inlet air evaporative cooling & supplementary firing for flat output at ambients above ISO (water-cooled, wet
mechanical draft cooling tower)
Figure 13 shows results similar to those of Figure 12, except that the gas turbine is now provided with an
evaporative inlet cooler, lessening the extent of supplementary firing needed on hot days. Comparing the
figures, we can see that at the hottest ambient, the evaporative cooler improves the gas turbine output
from about 80% to about 85% of nominal. The shortfall to be made up by the steam turbine is less, so we
only need the steam turbine to generate 28% additional power, not 40%. In this example, at the hottest
day, the duct burner exit temperature is a more manageable 830 °C (1526 °F), if it is at the boiler inlet.
16.3
EFFECT OF AMBIENT HUMIDITY ON COMBINED CYCLES AT FULL LOAD
The effect of ambient relative humidity on the gas turbine, on its own, was discussed in detail in §6.4.
The effect on the combined cycle may be strong, especially at high ambient temperatures, if the plant uses
evaporative cooling, either in wet cooling towers, or for the gas turbine inlet air, or both. Figure 14 shows
how substantial this effect can be for a combined cycle using both forms of evaporative cooling at a hot
ambient of 35 °C (95 °F). Under these conditions, a 20%-humid ambient can add 6% to output and
improve heat rate by 2% relative to a 60%-humid ambient. Conversely, a 100%-humid ambient causes a
similar degradation in plant performance.
16-12
 M. A. Elmasri, 1990-2002
Chapter 16: CC Off-Design Behaviour
Effect of Ambient Humidity on Typical 3PRH CC @ Tamb = 35 C (95 F)
WCC/MDCT/Evap Inlet Cooler
Power & Heat Rate, % of Values @ 60% RH
110
GT Power
108
ST Power
106
GT Heat Rate
104
CC Net Heat rate
CC Net Output
ST Pseudo Heat Rate
102
100
98
96
94
92
0
10
20
30
40
50
60
70
80
90
100
Ambient Relative Humidity, %
Figure 14. Power and heat rate corrections for ambient humidity on a hot day, for a combined cycle relying on
evaporative cooling both for the GT inlet air and for its wet mechanical draft cooling tower
If no evaporative cooling of any sort is used, or if the ambient is cold, the effect of relative humidity is
quite weak, and high humidity enhances output, both of the gas turbine, as discussed in §6.4, and of the
steam cycle, due to the greater specific heat of the exhaust gases. Figure 15 shows an example, with no
evaporative cooling and no wet cooling tower. In this example, it was assumed that the gas turbine firing
temperature control acts on exhaust temperature with an adjustment for compressor discharge pressure,
and does not compensate for ambient humidity. This results in a T4 that is essentially constant at different
humidities, and a declining T3 as humidity increases. If the gas turbine controls included humidity
compensation, to keep T3 constant as inlet air humidity increased, T4 would rise, and the performance
improvement with humidity would be stronger, both for the gas turbine and for the steam cycle.
Effect of Ambient Humidity on Typical 3PRH CC @ Tamb = 35 C (95 F)
Open Loop Water Cooled Condenser, No Evap Inlet Cooler
101
Power & Heat Rate, % of Values @ 60% RH
GT Power
ST Power
CC Net Output
GT Heat Rate
100.5
ST Pseudo Heat Rate
CC Net Heat rate
100
99.5
99
0
10
20
30
40
50
60
70
80
90
100
Ambient Relative Humidity, %
Figure 15. Power and heat rate corrections for ambient humidity on a hot day, for a combined cycle which does not
use any evaporative cooling
16-13
 M. A. Elmasri, 1990-2002
Chapter 16: CC Off-Design Behaviour
16.4
BEHAVIOUR OF COMBINED CYCLES AT DIFFERENT LOADS
To reduce output from an unfired combined cycle, gas turbine load is reduced. To obtain more output,
supplementary firing may be used. Figure 16 shows how the gas turbine’s and steam turbine’s
contributions to total output vary with changes in total, net combined cycle output. It also shows the
variation in plant net heat rate with plant load, and the variation of gas turbine and steam turbine pseudo
heat rates. These pseudo heat rates are defined in terms of the power output from each turbine and the
total plant fuel input, to both the gas turbine and duct burner, if used. Figure 17 helps to understand the
results, by showing the variation in the key measures of steam bottoming cycle efficiency.
Performance at varying load for typical 3PRH CC @ ISO conditions
Power & Heat Rate, % of nominal unfired
140
GT Power
130
ST Power
GT Pseudo Heat Rate
ST Pseudo Heat Rate
120
CC Net Heat rate
110
100
90
80
70
60
65
70
75
80
85
90
95
100
105
110
115
CC Net Output, % of nominal unfired
Figure 16. Power and heat rate at varying load. 100% refers to nominal, unfired condition
Performance at varying load for typical 3PRH CC @ ISO conditions
43
90
42
89
88
87
40
HRB efficiency, %
39
Steam cycle overall efficiency, %
Steam cycle internal efficiency, %
86
38
37
36
35
HRB efficiency, %
Steam cycle efficiencies, %
41
34
33
32
31
65
70
75
80
85
90
95
100
105
110
115
CC Net Output, % of nominal unfired
Figure 17. Steam bottoming cycle efficiencies at various loads. 100% refers to nominal, unfired condition
In both figures 16 and 17, 100% on the x-axis represents the output of the combined cycle with the gas
turbine at full-load, and without supplementary firing. In Figure 16, 100% on the y-axis also represents
the performance of the combined cycle with the gas turbine at full-load and the duct burner off. This is
assumed to be the nominal design condition of the triple-pressure reheat combined cycle plant, with one
16-14
 M. A. Elmasri, 1990-2002
Chapter 16: CC Off-Design Behaviour
60 Hz “F-class” gas turbine and one steam turbine, and with a water-cooled condenser and wet,
mechanical draft cooling tower. The plant’s 100% net output corresponds to 260 MW and its 100% net
heat rate corresponds to 6564 kJ/kWh (54.85% thermal efficiency).
In Figures 16 and 17, the curves end at a level of supplementary firing of 760 °C (1400 °F), with the duct
burner at the inlet to the heat recovery boiler, a reasonable level that requires no special measures in
boiler design. This level of supplementary firing allows roughly 14% more output than from an unfired
combined cycle of the same design.
The gas turbine model assumed for Figures 16 & 17 has an exhaust temperature which climbs as load is
reduced by closing its inlet guide vanes. The reduction in flow rate and the increase in temperature of the
exhaust gases both work to improve HRB efficiency as load is reduced, as seen in Figure 17. Also, the
increase in exhaust gas temperature increases the proportion of the total steam production at the high
pressure, improving steam cycle internal efficiency. The result is an increase in bottoming cycle overall
efficiency as GT load is reduced, as seen in Figure 17. Thus, the decrease in steam turbine output is less
than proportional to the decrease in gas turbine output, as seen in Figure 16. As gas turbine load is
reduced, its efficiency decreases, and, in this example, overwhelms the improvement in bottoming cycle
efficiency, causing net plant efficiency to fall, albeit slightly, as seen by the slight rise in net plant heat
rate in Figure 16.
When the gas turbine is at full-load and supplementary firing is instituted to increase output, the
bottoming cycle efficiency improves. This is due to higher heat recovery efficiency, as well as to the
increasing proportion of this heat recovery which goes to the high pressure steam, thereby improving
steam cycle internal efficiency. However, the increase in plant fuel input overwhelms the improvement in
bottoming cycle efficiency, and results in a falling net plant efficiency with supplementary firing level, as
seen in Figure 16. The pseudo heat rates for the gas and steam turbine in this diagram are defined in
terms of the output from each and the total plant fuel input. Whereas steam turbine output increases with
supplementary firing, as a proportion of the total fuel input, the gas turbine output falls as a percentage of
total fuel input, since it gets no benefit whatsoever from the duct burner fuel.
In the example of Figures 16 & 17, it was assumed that the steam turbine was sized to run with its throttle
valve wide-open at the nominal design point. The throttle pressure selected for nominal design, based on
optimising performance at this unfired condition, was 115 bars (1670 psia). As GT load is reduced and
steam production falls, the throttle valve was assumed to be left wide open, and the pressure left to slide,
in unison with the falling flow rate. With the gas turbine at full load, increasing supplementary firing
leads to higher steam production, and forces the pressure at the throttle to climb as more steam is forced
into the now undersized steam turbine inlet nozzles. This behaviour is depicted as “Design A” in Figure
18. At the maximum supplementary firing level, throttle pressure is 140% of nominal in our example, i.e.
161 bars (2335 psia). Thus, if the plant is designed according to these criteria, it must be capable of this
pressure, in its boiler, feedpumps and piping, as well as in its steam turbine. If the maximum
supplementary firing is used for a small percentage of the year, the extra costs may be hard to justify.
An alternative steam design would have oversized first stage nozzles at the nominal design-point, such
that when supplementary firing is at its full extent, the additional steam flow may be accommodated
without excessive pressure. Designs “B” and “C” in Figure 18 have steam turbines sized to accommodate
the maximum flow at full supplementary firing with an overpressure of just 10%, compared to nominal.
Both have the same steam turbine sizing, and the difference between them is in operating controls.
Design B begins to close its throttle valve only if the sliding pressure falls to 75% of nominal, but Design
C begins to close it if the sliding pressure falls to nominal.
16-15
 M. A. Elmasri, 1990-2002
Chapter 16: CC Off-Design Behaviour
Throttle pressures at varying load for three alternate designs
HPT Throttle Pressure, % of nominal unfired
150
Design A
140
Design B
Design C
130
120
110
100
90
80
70
65
70
75
80
85
90
95
100
105
110
115
CC Net Output, % of nominal unfired
Figure 18. Alternative designs to deal with the problem of steam turbine overpressure with supplementary firing
Figure 19 shows the implications for plant heat rate, expressed as a percentage of the nominal heat rate of
Design A, with its gas turbine at full-load and its duct burner off. Designs B and C are less efficient, due
to their lower pressures and to their throttling losses. This loss of efficiency must be traded off against the
additional capital costs of Design A, to find which method offers the best overall economics.
Plant Net Heat Rate at varying loads for three alternate designs
108
Design A
Plant Net HR, % of nominal unfired
107
Design B
Design C
106
105
104
103
102
101
100
65
70
75
80
85
90
95
100
105
110
115
CC Net Output, % of nominal unfired
Figure 19. Heat rate variation with load for alternative designs to deal with the problem of steam turbine overpressure
with supplementary firing
Before closing, it is worthwhile to mention that combined cycle heat rate may improve at loads slightly
below nominal design, depending on the gas turbine part-load characteristics. Figure 20 illustrates this by
a simulation of dual-pressure combined cycles based on three different types of gas turbines. The
machines assumed in these simulations are all between 40 and 65 MW. For the GT model in which part
load control is by IGV closure at constant T3, followed by T3 reduction after full IGV closure, combined
cycle heat rate improves as load is reduced, and is best in the load range 90-95%. Figure 12 of Chapter 7
illustrates the part-load exhaust characteristics of this sort of gas turbine. For the machine in which T4 is
held constant as load is reduced, by simultaneous closure of the IGV and reduction of T3, heat rate begins
to degrade as load is reduced from 100%. Figure 13 of Chapter 7 illustrates the part-load exhaust
characteristics of this type of gas turbine. Likewise for the aeroderivative in which both exhaust flow and
16-16
 M. A. Elmasri, 1990-2002
Chapter 16: CC Off-Design Behaviour
exhaust temperature fall together as load is reduced, with exhaust characteristics of the sort shown in
Figure 14 of Chapter 7.
Typical 2PCC Net Heat Rates at Part Load for Various GT Types
Net CC Heat Rate, % of Full Load
112
Heavy Duty-IGV followed by TIT
Heavy Duty, Constant TET
Aeroderivative
110
108
106
104
102
100
98
60
70
80
90
100
Net CC Output %
Figure 20. Heat rate at part-load for dual-pressure combined cycles based on gas turbines with different part-load
control characteristics
16-17
 M. A. Elmasri, 1990-2007
Chapter 17: Cost-Efficiency Tradeoffs
COST-EFFICIENCY TRADEOFF
Content Revised November, 2002. Updated November, 2007
© Maher Elmasri 1990-2007
17.1
COST OF GENERATING ELECTRICITY
The cost of producing electricity consists of two components, one proportional to the capital cost of the
plant and one proportional to the operating cost of the plant.
The first component, proportional to capital cost, consists of several sub-components:
(a) Debt repayment, both principal and interest to pay the banks and bondholders who financed the
plant construction. Assume this annual cost is Ra times plant capital cost. Typically, Ra = 0.07, say.
(b) Plant maintenance, overhauls and upkeep, salaries of plant personnel, insurance, property taxes, etc.
Whilst acknowledging that some of these costs, especially maintenance, are not solely a function of
original equipment cost, but depend also on operating hours, number of starts, etc, but in the interest
of simplicity we shall adopt the approximation of their proportionality to equipment capital cost,
which, at least, is true for plants that have similar operating hours and duty cycles. Assume this
annual cost is Rb times the plant capital cost. Typically, Rb = 0.06, say.
(c) Return on investment to the plant owners, who need to garner a certain, reasonable profit for having
caused the construction of the plant. Assume this annual cost is Rc times the plant capital cost.
Typically, Rc = 0.04, say.
Thus, the first component is an annual cost of R times the capital cost, where R is the fixed charge rate. For
typical situations, R = (Ra + Rb + Rc) should be in the range 0.15-0.20, and in the discussion given above it
sums up to 0.17.
The second component of electricity cost is fuel. If the cost of fuel per unit of its thermal energy input is H,
and the plant efficiency is η, then the cost of fuel to generate a unit of electrical energy is (H/η).
Thus the cost of electricity, E, in $ per kWh is
E=
CR H
+
T
η
……………. (1)
where
C
is the capital cost in $/kW of capacity
R
is the fixed charge rate, the sum of (a)-(c) above, which in our example adds up to 0.17
T
is the number of full-load-equivalent operating hours per year
H
is the price of fuel thermal energy, per kWh of fuel LHV energy content
η
is the plant efficiency on an LHV basis.
In the present analysis, the fixed charge factor R is used to annualize capital cost recovery and also to
include all the non-fuel components of plant operating cost. In many economic analyses, the fixed charge
rate is used only for the annualized capital cost recovery, in which cases its value is normally smaller, say
0.09-0.12. Because it includes all non-fuel costs in the present analysis, a reasonable value for it is 0.150.20. Some may argue that the non-fuel operating costs are not truly proportional to capital cost, and hence
cannot be rolled into R. The author recognizes this viewpoint, but believes that the approximation of
assuming that non-fuel costs are proportional to capital cost is reasonable, at least for plants that are similar
12-1
 M. A. Elmasri, 1990-2007
Chapter 17: Cost-Efficiency Tradeoffs
in type and in duty cycle. The simplicity and lucidity of the resulting conclusions outweighs the error
perceived in this approximation.
We now define a dimensionless fuel price f, as fuel price/electricity price in consistent units,
f =
H
E
……………….
(2)
where
H
is the price of fuel thermal energy, $ per kWh of fuel LHV energy content
E
is the price of electricity, $/kWh
0.55
0.5
Dimensionless fuel price (LHV), f
0.45
0.4
0.35
Fuel Price
$/MMBTU
(HHV)
0.3
$12
$11
$10
0.25
$9
$8
0.2
0.15
0.1
$/MMBTU $/MMBTU
HHV
LHV
Figure 1. The dimensionless fuel price “f” for a range of
$7
$6
$5
$4
electricity
$3
and
12
11
10
9
8
7
6
5
4
fuel
3
13.32
12.21
11.10
9.99
8.88
7.77
6.66
5.55
4.44
prices
3.33
$/kWh
LHV
$/GJ
HHV
$/GJ
LHV
0.0455
0.0417
0.0379
0.0341
0.0303
0.0265
0.0227
0.0189
0.0152
0.0114
14.01
12.85
11.68
10.51
9.34
8.18
7.01
5.84
4.67
3.50
12.63
11.57
10.52
9.47
8.42
7.36
6.31
5.26
4.21
3.16
With 0.05
price of fuel, one has to be clear on whether the cost per unit energy is based on the LHV or HHV
0.04the0.05
0.07 not
0.08 matter
0.09 which
0.1
0.11
0.13 as
0.14long
0.15as the plant efficiency is defined on the
energy of
fuel.0.06It does
basis0.12is used,
Electricity Price $/kWh
same basis. However, since we
have adopted the LHV as the basis for all our discussions, then we shall
base both the efficiency and the fuel price on LHV.
The cost of natural gas fuel in the US is normally quoted in $/MMBTU• based on HHV. Thus, to calculate
H in $ per kWh LHV, we must first adjust fuel cost to an LHV basis, then convert from BTU to kWh.
Figure 1 shows the values of f corresponding to a range of electricity and fuel prices defined in the US
traditional way of $/MMBTU(HHV). A table is shown alongside the figure showing the corresponding fuel
price in various units, US and SI. It is assumed that the HHV/LHV ratio is 1.11, typical for natural gas that
is mostly methane.
From Fig. 1, we see, for example, that if electricity were 6¢/kWh and fuel $4/MMBTU, then f ≈ 0.3. If
electricity were 7.5¢/kWh and fuel $6/MMBTU, or electricity were 10¢/kWh and fuel $8/MMBTUthen f is
still ≈ 0.3. As another example, if electricity were 6¢/kWh and fuel $6/MMBTU, then f ≈ 0.38, the same as
with electricity at 8¢/kWh and fuel at $8/MMBTU. The “normal” range of f is 0.25 - 0.45.
To complete our simple economic analysis, we introduce the capacity factor ψ
“MM” is a “classical” abbreviation for million, so MMBTU stands for a million BTU. “M” is also used as an
abbreviation for thousand, so MBTU stands for a thousand BTU. These abbreviations are becoming archaic and are
not even consistent with Latin numerals, where “MM” would actually imply two thousand and not one million.
•
12-2
 M. A. Elmasri, 1990-2007
Chapter 17: Cost-Efficiency Tradeoffs
ψ =
T
8760
…………….
(3)
Since a year consists of 8760 hours, a plant that operates full-load for T hours will have a capacity factor of
T/8760. Capacity factor as used here represents the portion of the year of equivalent full-load operation of
the power plant.
Now equation (1) can be rearranged to give the justifiable capital cost of a plant in $/kW of capacity as:
C=
8760ψE 
f
1 −

R  η



……………. (4)
As a simple numerical illustration, if we assume ψ=0.5, E=0.06 $/kWh, R=0.17, f=0.25, η=0.5, we get a
justifiable capital cost of C=$773/kW for such a combined cycle. This makes sense. If, on the other hand,
fuel was more expensive relative to electricity, with f = 0.35 say, the justifiable capital cost would be only
$464/kW. Since it is unlikely that a 50%-efficient plant can be built at such a low cost, this means that with
f = 0.35, either the capacity factor ψ would need to be higher than 0.5, or the fixed charge rate R lower than
0.17, or the electricity more expensive than 6¢/kWh (with proportionately higher fuel price, since f = 0.35).
From equation (4), if f = η, then C would equal zero. Naturally, if the ratio of fuel price to electricity price
were equal to the plant efficiency, then the revenue from electricity would just pay for the fuel, leaving no
money to pay for any of the capital-related costs, so its justifiable capital cost would be zero. To the extent
that η exceeds f, the quantity in the parentheses of equation 4 becomes larger. In other words, as η exceeds
f, the electricity produced has more value than the fuel needed to produce it, and the justifiable capital
investment in the plant increases.
17.2
COMPARING CAPITAL COST
DESIGNS
AND
EFFICIENCY
FOR
ALTERNATIVE PLANT
Let us assume that a decision has been made to build a natural gas fired power plant at a certain site and for
a certain anticipated duty. Presumably, this decision would have been made after a far more complex, and
correspondingly far more accurate, analysis than given by equation (4) above.
The next phase is to compare different plant design options to see which is economically superior. For this
phase, equation (4) can provide valuable insight. The various design options each have a different capital
cost and a different efficiency, so we write equation 4 for the base case plant design, identifying its capital
cost C and its efficiency η with the subscript 0,
C0 =
8760ψE 
f 

1
−
R  η 0 
…………. (5)
then we write it for an alternative plant design, which we identify by having no subscript on C and η, i.e. it is
the same equation (4). We then divide both sides of equation 4 by those of equation 5, so we find that the
quantity
8760ψE
cancels out, rather conveniently, giving just
R
1 − f 
η
C
= 
C0  1 − f 

η 0 

…………….
12-3
(6)
 M. A. Elmasri, 1990-2007
Chapter 17: Cost-Efficiency Tradeoffs
Equation (6) is a very useful tool when comparing alternative designs on an equal footing. It is remarkably
simple, and eliminates superfluous factors that may complicate the comparison. For instance, experienced
people may differ as to which value of R to use, or what capacity factor ψ to assume for the new plant, but
these cancel out in this comparative approach, so their precise value does not even matter. To avoid any
misunderstanding, it is worth emphasizing that the values of R and ψ do matter, and are very important at
the previous phase of the development, when the decision is being made whether or not to build a plant of a
particular type at a particular site. Once this decision has been made, and we are at the stage of comparing
design alternatives for the plant, the values of R and ψ no longer matter as seen in equation 6. The only
parameter affecting the selection between alternative designs, other than the cost and efficiency of each
design, is f, the dimensionless fuel price,
To illustrate the application of equation 6, let us say we have a base case design of a combined cycle of 55%
net thermal efficiency. Such a plant is likely to be of a triple-pressure reheat configuration. We wish to find
out if alternative designs that have a lower capital cost but lower efficiency, or higher capital cost but higher
efficiency, are better or worse than our base case. To do this, we plot equation (6) with η0 = 0.55 for various
values of f as shown in Figure 2.
As an example of using Fig. 2 to compare plant options, let us suppose that fuel were cheap relative to
electricity, with f = 0.2. That curve at a 40%-efficient alternative plant shows a (C/C0) of 0.79. This means
that if the 40%-efficient alternative plant were to cost 79% of the cost of the reference plant, it will have the
same economics. If the 40%-efficient alternative were to cost less, it would be economically superior to the
55%-efficient reference plant, and vise versa. Now, if a 55%-efficient combined cycle costs $700/kW, and a
40%-efficient aeroderivative gas turbine plant costs $400/kW, i.e. less than (0.79 x 700 = 553), then the
40%-efficient aeroderivative plant is superior economically to the 55%-efficient reference combined cycle
under these circumstances.
Justifiable capital costs of alternative plant designs for the same duty cycle as a
55% -efficient base case design
1.4
f=0.20
Justifiable Cost, dimensionless C/Co
1.3
f=0.25
1.2
f=0.30
1.1
f=0.35
f=0.40
1.0
f=0.45
0.9
0.8
0.7
0.6
0.5
0.4
0.3
0.2
0.1
0.0
0.34
0.36
0.38
0.40
0.42
0.44
0.46
0.48
0.50
0.52
0.54
Plant Efficiency
Figure 2. Plot of equation 6 with a base case of η0 = 0.55
12-4
0.56
0.58
0.60
 M. A. Elmasri, 1990-2007
Chapter 17: Cost-Efficiency Tradeoffs
Now, if fuel were slightly more expensive relative to electricity, with f = 0.25, the 40%-efficient alternative
shows a (C/C0) of 0.69. Using the figures of the previous paragraph, the 40%-efficient aeroderivative gas
turbine would need to cost less than 0.69 x 700 = $483/kW to beat the 55%-efficient reference combined
cycle at $700/kW. This is still likely in practice.
Moving on to more expensive fuel relative to electricity, with f = 0.30, the 40%-efficient alternative plant
shows a (C/C0) of 0.55. Using the earlier figures, the 40%-efficient aeroderivative gas turbine would need to
cost less than 0.55 x 700 = $385/kW to beat the 55%-efficient reference combined cycle at $700/kW. This
is unlikely, so the combined cycle wins.
Considering even more expensive fuel relative to electricity, with f = 0.35, the 40%-efficient alternative
plant shows a (C/C0) of 0.34. Using the earlier figures, the 40%-efficient aeroderivative gas turbine would
need to cost less than 0.34 x 700 = $238/kW to beat the 55%-efficient reference combined cycle at
$700/kW. Thus the reference plant wins handily.
Considering even more expensive fuel relative to electricity, with f = 0.40, the 40%-efficient alternative
plant shows a (C/C0) of zero, naturally, so it would have to be free!
Moving to the other end of the spectrum, we can see that with expensive fuel relative to electricity, f=0.45, a
60%-efficient alternative, perhaps an advanced combined cycle with a steam-cooled gas turbine, has a
(C/C0) of 1.375. Thus, if it could be built for under 1.375x700 = $963/kW, it would be economically
superior to our 55%-efficient reference combined cycle at $700/kW. With f = 0.4, it needs to cost less than
1.22 times the reference plant; and with f = 0.35, it needs to cost less than 1.15 times the reference plant.
In concluding this section, it is again worth emphasizing that the above comparisons are all for different
plant options that will operate on the same duty cycle.
It is also worth emphasizing that f is a ratio. Low f can be due to either cheap fuel or expensive electricity.
Indeed, even with expensive fuel, if one is considering a peaking plant, to run only at times when electricity
is very expensive, then f is still low, and, as the example above showed for low f, the aeroderivative option
beats the combined cycle. Likewise, high f can be due to either expensive fuel or cheap electricity, and this
favours the more expensive and more efficient options.
17.3
TRADEOFF BETWEEN SPECIFIC CAPITAL COST & EFFICIENCY
17.3.1 THE DIFFERENTIAL FORM OF THE COST-EFFICIENCY TRADEOFF EQUATION
Equation 6 is useful for a wide-angle view of the cost-efficiency tradeoff. To zoom in and study small
changes, it is useful to take its slope, the partial derivative of dimensionless specific cost (C/C0) with respect
to relative efficiency (η/η0). This gives equation (7) below, which is also plotted in Fig. 3.
∂ C 
 C0  = f
∂ η  η0 − f
 η0 
…………….
(7)
Justifiable slope of (C/Co) with respect to (η/ηo)
5.0
4.5
4.0
3.5
∂ C 
 C0 
∂ η 
 η0 
3.0
2.5
f=0.45
f=0.40
f=0.35
f=0.30
f=0.25
f=0.20
12-5
 M. A. Elmasri, 1990-2007
Chapter 17: Cost-Efficiency Tradeoffs
Figure 3. Plot of equation 7 with base cases ranging from η0 = 0.53 to η0 = 0.59
To illustrate the use of Figure 3 (and equation 7), let us assume that f = 0.38 (corresponding to 7¢/kWh
electricity and $7/MMBTU fuel, say). Now, suppose our base case design is a 55%-efficient combined
cycle that has a capital cost of $700/kW. Figure 3 (or equation 7) shows that for η0 =0.55 and f = 0.38, the
slope of justifiable dimensionless specific cost relative to efficiency is 2.24, meaning that a 1% efficiency
improvement is worth a 2.24% increase in $/kW. This means that if we “tweaked” the design, such as, for
instance, by using smaller boiler pinch points, and increased the efficiency by 1% (from 55% to 1.01 x 55%
= 55.55%), the plant would be worth 2.24% more per kW, i.e. 1.024 x $700/kW = $717 $/kW. If such a
“tweak” increased the specific cost to less than $717/kW, then it is worthwhile, but if it increased the
specific cost to more than $717/kW, then we are better off without it.
These tradeoffs are very sensitive to f. For instance, at 55% base efficiency, if f were 0.3, the slope of
justifiable dimensionless specific cost relative to efficiency is just 1.2; with f = 0.35, it is 1.75; with f = 0.40,
it is 2.67; and with f = 0.45, it is 4.5. Naturally, with f = η0 , it is infinity.
Fuel prices have been volatile, as shown by Figure 4. The dimensionless fuel price, f, as used in the analysis
should be less volatile than fuel price alone, as increases in fuel price are usually accompanied by increases
in electricity prices. In the absence of more detailed analysis, the author suggests using f=0.4, which, for a
typical modern combined cycle corresponds to a slope of justifiable dimensionless specific cost relative to
efficiency of 2.5 as seen from Fig. 3.
Figure 4. Natural gas prices in the US 204-2007. The lowest curve is for electric power. The y-axis is price per
thousand cubic feet, which is about the same as the price per MMBTU (Source: US Department of Energy EIA)
17.3.2 EXAMPLE ON OPTIMISING AIR-COOLED CONDENSER PRESSURE
The above arguments, leading to equation 7, are based on straightforward math telling us how much extra
specific cost is justified to improve a plant’s efficiency. To use the results in optimising a plant’s parameters
12-6
 M. A. Elmasri, 1990-2007
Chapter 17: Cost-Efficiency Tradeoffs
requires a method of calculating how much a real plant’s efficiency and specific cost will change when we
change its design. Calculating a change in efficiency can be done with reasonable certainty using programs
such as GT PRO, but calculating the likely change in cost is, by its nature, less accurate. So, with this
caveat in mind, we shall present an example of a combined cycle with an air-cooled condenser and seek its
optimum condenser pressure. The combined cycle is a typical 3-pressure reheat plant at a 25 ºC design
ambient, and its heat balance is illustrated in Fig. 5.
Gross Power
Net Power
Aux. & Losses
LHV Gross Heat Rate
LHV Net Heat Rate
LHV Gross Electric Eff.
LHV Net Electric Eff.
Fuel LHV Input
Fuel HHV Input
Net Process Heat
263962 kW
255399 kW
8563 kW
6404 kJ/kWh
6619 kJ/kWh
56.22 %
54.39 %
469553 kWth
521023 kWth
0 kWth
Ambient
1.013 P
25 T
60% RH
115 p
565 T
58.84 M
Stop Valve
93952 kW
Cold Reheat
31.52 p
375.5 T
57.09 M
27.5 p
565 T
63.82 M
0.1073 p
47.22 T
71.92 M
4.005 p
262.2 T
6.461 M
0.0774 M
Hot Reheat
1.013 p
99.7 T
428.6 M
to HRSG
LP
HP
IP
LPB
4.5 p
147.9 T
7.916 M
IPB
31.67 p
236.9 T
6.733 M
HPB
120.8 p
325.2 T
58.84 M
194.4 T
157.9 T
275 T
249.9 T
484.8 T
338.2 T
47.26 T
72.11 M
p [bar] T [C] M [kg/s], Steam Properties: Thermoflow - STQUIK
0 10 25 2007 10 44 19 fil C \S i
2007\S
10 07\S 17 C t Effi i
CH4
9.382 M
469553 kWth LHV
1.013 p
25 T
419.3 M
1.003 p
25 T
419.3 M
1.044 p
640.6 T
428.6 M
GE 7251FB
@ 100% load
170010 kW
GT PRO 17.0 Maher Elmasri
\2007 3PRH BASEACC GTP
Figure 5. Heat balance of the combined cycle used for an example of optimising condenser pressure
We shall try eight alternative designs, each with a different condenser pressure and calculate the
performance of each design, as well as estimate its capital cost with the GT PRO/PEACE software system.
The capital cost increases as condenser pressure is reduced mostly due to the increased condenser surface
area, but in the GT PRO/PEACE design system, the steam turbine’s exhaust end will be re-sized, also
increasing cost; and various other minor differences in the design of many components will also affect the
cost. Naturally, lowering condenser pressure also increases net power output, so the increases in specific
cost are less than proportional to the increases in total cost. Table 1 below summarises the results.
Table 1. Summary of pertinent results for the alternative design cases of the combined cycle of Fig. 5
Condenser pressure, bar
Specific Cost $/kW
Increase in specific cost over previous case, $/kW
D(C/Co) increase divided by previous specific cost
Plant net η
Increase in net η over previous case
D(η/ηo) increase divided by previous case η
D(C/Co) / D(η/ηo)
0.27
0.24
650.4702 652.3523
1.882068
0.002893
0.528067 0.530448
0.002381
0.004509
0.642
12-7
0.21
654.6482
2.29584
0.003519
0.533136
0.002688
0.005067
0.695
0.18
659.2374
4.589251
0.00701
0.535578
0.002443
0.004582
1.530
0.15
666.2567
7.019299
0.010648
0.538528
0.00295
0.005507
1.933
0.12
675.601
9.344345
0.014025
0.54221
0.003682
0.006838
2.051
0.09
697.9332
22.33219
0.033055
0.54634
0.004129
0.007616
4.340
0.06
761.0166
63.08337
0.090386
0.550369
0.004029
0.007375
12.255
 M. A. Elmasri, 1990-2007
Chapter 17: Cost-Efficiency Tradeoffs
If we follow the recommendation of 2.5 for the slope of justifiable dimensionless specific cost relative to
efficiency, we see from the bottom line of Table 1 that this is attained at a condenser pressure of between
0.12 and 0.09 bars, at about 0.11 bars, say.
From the bottom line of Table 1, we can also glean the optimum selection of condenser pressure if fuel were
cheaper. For example, if f were 0.3, say, Fig. 3 tells us that the justifiable slope is about 1.2, which would
point us towards a design condenser pressure of about 0.19 bars.
A word of caution is due. When undertaking an analysis of the sort summarised in Table 1, we are often
dealing with very small differences, particularly in specific cost. A design change to improve performance,
such as reducing condenser pressure, leads to more power output and higher cost. Their ratio, the $/kW may
change very little. Also the efficiency changes are very small. Thus, the bottom line of Table 1 is the ratio
of two very small numbers. Since zero divided by zero is indeterminate, we must make sure that the two
small numbers we are dividing have a large number of significant digits and adequate accuracy.
17.3.3 THE VALUE OF FUEL-FREE INCREMENTAL OUTPUT AND ITS USE IN OPTIMISATION
When optimising a combined cycle, one frequently makes design changes that increase output without
increasing fuel input. One example is lowering condenser pressure, as discussed in §17.3.2. Another
example is reducing pinch points. Yet another is changing the design from a dual-pressure to a triplepressure cycle. As a matter of fact, most “tweaks” of a combined cycle to improve its output do not change
fuel input. In these cases, the percentage increase in efficiency is the same as the percentage increase in
output.
Suppose we have a power plant with a capacity of P0 kW, an efficiency of η0 , and a capital cost of C0 $/kW.
Its value is (P0 C0) dollars. Now we introduce a design change that increases capacity by ΔP kW and
efficiency by Δη. The increase in value of the plant in dollars is
∆$ = ∆PC0 + P0  ∂C  ∆η …… (8)
 ∂η  just
where the first term is the value of the increased capacity at the original efficiency, and the second term
represents the value of the justifiable increase of specific cost applied it to the entire capacity, which has
become more efficient.
If the design change increases power without affecting fuel consumption, then the efficiency and power
increase are proportionate
∆η
η0
=
∆PFF
…………………………….
P0
(9)
where the subscript “FF” has been added to ΔPFF to indicate that the extra power is fuel-free for this
equation to be true. Thus, equation 8 becomes
η
∆$ = ∆PFF C0 + P0  ∂C  ∆PFF 0 ……………. (10)
η
∂

 just
P0
Rearranging and simplifying, equation 10 gives

C 
 ∂ C0  
∆$
= C0 1 +
 ……………………….
η
∆PFF


 ∂

 η0  

which, in view of equation 7 now yields
12-8
(11)
 M. A. Elmasri, 1990-2007
Chapter 17: Cost-Efficiency Tradeoffs

 ∆$
 ∆P 
f
FF 

= 1+
C0
η0 − f
…………….
(12)
Equation 12 is plotted in Fig. 6 to show the value of adding fuel-free incremental capacity as a multiple of
the value of the original capacity. With f = 0.35, fuel-free incremental capacity is worth about 2.7 times as
much as the specific cost, and with f = 0.4., fuel-free incremental capacity is worth about 3.5 times as much
as the specific cost. This is the suggested multiple in the absence of a more detailed analysis.
The value of gaining a fuel-free kW as a multiple of the original specific cost Co
6.0
Value of incremental fuel-free kW / Co
5.5
5.0
4.5
f=0.45
f=0.40
f=0.35
f=0.30
f=0.25
f=0.20
4.0
3.5
3.0
2.5
2.0
1.5
1.0
0.53
0.54
0.55
0.56
0.57
0.58
0.59
Base Plant Efficiency
Figure 6. Plot of equation 12 with base cases ranging from η0 = 0.53 to η0 = 0.59
17.3.4 THE EXAMPLE ON OPTIMISING AIR-COOLED CONDENSER PRESSURE WITH THE METHOD
OF FUEL-FREE CAPACITY VALUE
We now revisit the example of §17.3.2 to show that the same conclusion can be reached by considering the
cost of the extra fuel-free capacity. Table 2 shows the results of the same series of GT PRO/PEACE cases,
summarising the net output and total cost of each case, then looking at the cost of the incremental output in
$/kW gained in each design over the previous case, and its ratio to the capital $/kW of the previous case.
Table 2. Summary of pertinent results for the alternative design cases of the combined cycle of Fig. 5
Condenser pressure, bar
Specific Cost, Co $/kW
Plant net output, kW
Increase in net output over previous case, kW
Plant total cost, k$
Increase in total cost over previous case, k$
Cost of fuel-free incremental capacity, $/kW
Ratio (Incremental fuel free $/kW / previous Co $/kW)
0.27
0.24
0.21
0.18
0.15
650.4702 652.3523 654.6482 659.2374 666.2567
247956
249074
250336
251483
252868
1118
1262
1147
1385
161288
162484
163882
165787
168475
1196
1398
1905
2688
1069.8
1107.8
1660.9
1940.8
1.645
1.698
2.537
2.944
0.12
0.09
0.06
675.601 697.9332 761.0166
254597
256536
258428
1729
1939
1892
172006
179045
196668
3531
7039
17623
2042.2
3630.2
9314.5
3.065
5.373
13.346
As stated in §17.3.3, it is worthwhile to spend about 3.5 times as much as the base specific cost on fuel-free
incremental capacity. (This corresponds to the recommendation of 2.5 for the slope of justifiable
dimensionless specific cost relative to efficiency given in §17.3.1). The bottom line of Table 2 shows that
this ratio is less than 3 for all cases of condenser pressure above 0.12, i.e. that each case is preferable to its
12-9
 M. A. Elmasri, 1990-2007
Chapter 17: Cost-Efficiency Tradeoffs
predecessor, as it adds incremental fuel-free capacity at below its justifiable cost. The value of 3.5 is
attained at condenser pressure between 0.12 and 0.09, at about 0.11 bars, say.
The method of fuel-free incremental capacity gives the same result as the method of finding the slope of
justifiable dimensionless specific cost relative to efficiency, described in §17.3.1. Actually, the two methods
are the same, as evident by the mathematical derivations and the similarity of equation 12 to equation 7.
Equation 12 can be logically derived from equation 7 by inspection, but the math of its derivation was given
for completeness. However, in application, the method of fuel-free incremental capacity is preferable,
because (a) the concept is physically clear rather than mathematically arcane; and (b) it avoids taking
numerical derivatives as ratios of very small differences which can amplify rounding errors.
17.3.5 DIFFERENCE BETWEEN NPV AND IRR OPTIMISATION
In project financial evaluations, two alternative methods are commonly used:
Net Present Value (NPV): A discount rate is assumed, such as 12% or 15%. The net cash flow estimated
for each future year is discounted by this annual rate, according to how far into the future it will occur, to
find is "present value" (which is its reduced equivalent in today's money). The total present value of all the
future years' cash flows must be positive, and must exceed the project's initial investment. The extent by
which the total present value exceeds the project's initial investment is known as the Net Present Value
(NPV). It is the "value" created by the project in today's money.
Internal Rate of Return (IRR): Is the discount rate at which the NPV would be zero, which is effectively
the annual rate of return earned by investing in the project.
When a project is being optimised, one may seek the technical design choices that result in maximising
either the NPV or the IRR. The optimal choice of a technical parameter that maximises NPV is not the same
as that which maximises IRR. This is because NPV is dimensional, and IRR is dimensionless. For
example, assume a design change makes the plant produce 1% more power at the same efficiency, and also
increases its cost by 1%, it will increase its NPV by 1% but have no effect on its IRR. NPV increases
because the owner ends up with a bigger plant, a bigger investment, and a proportionately greater cash flow
in absolute dollars. On the other hand, the ratio of increased cash flow to increased capital investment does
not change, hence the IRR does not change.
The general rule is that optimising for NPV produces designs that are biased towards higher capital cost and
higher performance than optimising for IRR. The former is appropriate to maximise absolute profit if
availability of capital were unconstrained, whereas the latter is appropriate to maximise return in proportion
to the amount actually invested.
In the example of §17.3.2, the optimisation may just as well have been done directly by the GT
PRO/PEACE system, either on a Net Present Value (NPV) basis, or on a Return on Investment (IRR) basis.
The results are shown in Figure 7 using typical financial parameters and with electricity at $0.06/kWh and
fuel at $6/MMBTU. These correspond to f=0.38.
Fig. 7 shows that the optimum condenser pressure on an NPV basis is about 0.1 bar, whereas the optimum
on an IRR basis is about 0.12 bar. As expected, NPV is biased towards investing more capital and getting a
larger cash flow, favouring more “aggressive” designs, whereas IRR favours the ratio of income to capital,
rather than the absolute amount of cash flow.
12-10
 M. A. Elmasri, 1990-2007
Chapter 17: Cost-Efficiency Tradeoffs
Example on optimisation of an air-cooled condenser pressure,
illustrating difference between IRR and NPV optima
19.0
96,000
92,000
90,000
ROI
18.0
NPV
88,000
86,000
17.5
84,000
Net Present Value, k$
Return on Investment, %
94,000
18.5
82,000
17.0
80,000
16.5
78,000
0.27
0.24
0.21
0.18
0.15
0.12
0.09
0.06
Condenser pressure, bar
Figure 7. Example of optimisation of ACC design pressure on an NPV basis and on an IRR basis, directly via the GT
PRO/PEACE software system.
17.4
EXAMPLE OF COST-EFFICIENCY TRADE-OFF
CYCLE CONFIGURATIONS
FOR
DIFFERENT COMBINED
Figure 8. Cost-efficiency trade-off between the cycle configurations of Chapter 12.
As another example, the various combined cycle designs of Chapter 12 are compared in Figure 8. Absolute
costs in $ or € are avoided and only ratios are used based on the formulation presented in §17.3.2, equation
12. Absolute costs and currency conversion rates are volatile, and therefore the PEACE cost estimation
software used, or, for that matter, any other cost-estimation technique, is much more accurate when we
consider ratios, rather than absolute costs. This is the strength of the formulations given in this chapter.
12-11
 M. A. Elmasri, 1990-2007
Chapter 17: Cost-Efficiency Tradeoffs
From Fig. 8, one may draw the following conclusions:
The least expensive performance-enhancing options over the 1P base case are at ≈ 2.6-2.8 for the ratio
(Δ$/ΔPFF)/C0 . This corresponds to f ≈ 0.33. Thus, if fuel is cheap enough to result in f <≈ 0.33, the base
case single pressure case is more economical than any of the more efficient options.
If fuel were expensive enough, with f >≈ 0.33, a better cycle than the base 1P case is justified. Fig. 8 shows
that we can eliminate 1PRH, 2P, and 3P, all relative to 2PRH, since the latter offers the best efficiency
amongst these options, and the incremental cost for gaining its improved power and efficiency is lower than
any of the former. The 2P-PMRH option is not as easily dismissed. First, it is clear that 2P-PMRH is
superior to 2P, since it is both more efficient, and its incremental cost to gain its efficiency advantage is less.
Thus, if a less complex cycle than “true reheat” were sought, 2P-PMRH is distinctly superior to 2P. The
only caveat is that 2P-PMRH requires a large enough ST to have a cross-over between HP/IP and LP
casings to properly mix the hot IP steam, and, in typical practice, if the ST is large enough for a cross-over,
the gas turbines are usually large enough and their exhaust temperature high enough to employ “true reheat”
steam cycles. However, there may be situations where the fuel price is not high enough to justify the true
reheat cycles over the 2P-PMRH. Thus, a new comparison in which the 2P-PMRH is the base case is
needed to clarify the differences between it and the true reheat cycles. This is shown in Figure 9.
Figure 9. Cost-efficiency trade-off between the “true reheat” cycles and the PMRH configuration.
Fig. 9 shows that the 2P-HPRH is justifiable over the 2P-PMRH at just under 2.8 for the ratio (Δ$/ΔPFF)/C0 .
This corresponds to f ≈ 0.35. Thus, if the relative cost estimates were perfect, one would conclude that 2PPMRH is the preferred option when 0.33 ≈<f <≈ 0.35. Naturally, this band is too narrow to draw such a
conclusion at any moment in time for a plant that will likely operate for more than twenty years, and one
may safely conclude that true reheat is preferable, as it is more efficient.
It is also clear from Fig. 9 that 2P-HPRH is superior to 2PRH, since the former is more efficient and its
incremental cost to gain its efficiency and power advantage is also lower. Thus, 2PRH should be dismissed
except in situations where a relatively high stack temperature is sought, such as when fuel with some
sulphur is used for much of the plant’s operating life.
12-12
 M. A. Elmasri, 1990-2007
Chapter 17: Cost-Efficiency Tradeoffs
This leaves 3PRH and 2P-HPRH as the preferred true reheat options. To distinguish between them requires
a comparison in which the 2P-HPRH is the base case. This is shown in Figure 10, which suggests that the
extra cost of the 3PRH is only justified when the ratio (Δ$/ΔPFF)/C0 is ≈ 4.8, corresponding to f ≈ 0.43, a
very high fuel/electricity price ratio. Thus, it appears from this analysis that 2P-HPRH may be preferable to
3PRH. However, considering that the absolute differences between 2P-HPRH and 3PRH are small, the
slopes and gradients derived from such small differences may be of dubious reliability. Thus, it is more
appropriate to conclude that, perhaps, 3PRH is the preferred solution; particularly that (a) it is marginally
more efficient, and (b) it has been adopted as a standard, suggesting that economical analysis done by others
in the industry has shown it superior.
Figure 10. Cost-efficiency trade-off between the 3PRH and the 2P-HPRH configurations.
Thus, in summary:
(a) With cheap fuel/electricity, f <≈ 0.33, 1P is economically advantageous despite its low efficiency.
However, to adopt it, one must be confident that f will remain that low over the plant’s life span.
(b) With more expensive fuel/electricity, f >≈ 0.33, 3PRH or 2P-HPRH are roughly equivalent as the
best solutions, with 3PRH preferred at higher values of f.
(c) If a 2P configuration is contemplated for simplicity in a plant large enough to have a steam turbine
cross-over, 2P-PMRH is superior to 2P.
12-13
 M. A. Elmasri, 1990-2002
Chapter 18: Cogeneration System Selection
COGENERATION SYSTEM SELECTION
Content Revised May, 2002
© Maher Elmasri 1990-2002
18.1
CHOOSING A CONFIGURATION FOR COGENERATION WITH GAS TURBINES
18.1.1 RATIO OF A GAS TURBINE’S STEAM GENERATION POTENTIAL TO ITS POWER OUTPUT
A medium heavy duty gas turbine (such as a GE Frame 6F) converts about 34% of its fuel’s LHV energy to
electrical power. Of the remaining 66%, about 64.5%, emerges as exhaust heat and 1.5% as miscellaneous
losses (gearbox, mechanical and electrical). An unfired HRB can recover about 80% of the exhaust heat,
generating steam with an energy content (80% x 64.5%) ≈ 52% of the GT fuel’s LHV energy. Thus, the
ratio of potential steam generation to power generation is about 52/34 ≈ 1.5, without supplementary firing.
An advanced aero-derivative gas turbine (such as a GE LM6000 or Rolls-Royce Trent) converts about 41%
of GT fuel LHV energy to electricity. Of the remaining 59%, about 58%, emerges as exhaust heat and 1%
as miscellaneous losses (mechanical and electrical). Without supplementary firing, the modest exhaust
temperature of such a machine would normally allow roughly 70% of the exhaust heat to be recovered to
make steam. Thus, the steam produced would have an energy content that is roughly the same as the power
output, i.e. about 41% (.7 x .58 ≈ 0.41) of the GT Fuel LHV energy. The ratio of steam energy to power
generated is about 1, without supplementary firing.
Let us define a parameter σ as follows:
σ=
Energy _ of _ steam _ that _ can _ be _ generated _ from _ unfired _ GT _ exhaust
,
GT _ Power _ Output
which can be stated algebraically as
σ = η HRB
(1 − η −  misc )
η
............................ (1)
where
ηHRB is the heat recovery efficiency, defined in Chapter 9 §4.b
η is the GT efficiency
lmisc are the miscellaneous GT energy losses, such as GT mechanical loss, gearbox loss, generator
loss, etc. This is usually on the order of 0.01 for direct drive, and 0.015 with a gearbox, but
may be significantly larger if the GT rejects heat from its cooling air.
From the introductory discussion, σ ≈ 1.5 for a typical modern, medium heavy duty GT, and σ ≈ 1 for a
typical, modern aero-derivative. Although primarily dependent on the GT efficiency and exhaust
temperature, σ also depends on the number of HRB pressures and the steam and feedwater conditions, since
these parameters affect ηHRB. Table 1 shows values of σ for typical gas turbine models and three different
sets of steam parameters, chosen to be representative of cogeneration applications.
18-1
 M. A. Elmasri, 1990-2002
Chapter 18: Cogeneration System Selection
Table 1. Values of σ (unfired steam generation potential/electrical output) for some GT models
(ISO conditions, Feedwater @ 90 °C (194 °F) and typical inlet/exhaust losses & design assumptions)
GE Frame
6FA
Nominal GT efficiency, η %
Nominal GT Exhaust Temperature
σ for single-pressure, 20 bar/250 °C
(290 psia/482 °F)
σ for single-pressure, 50 bar/400 °C
(725 psia/742 °F)
σ for dual-pressure, 70 bar/435 °C &
7 bar/200 °C (1015 psia/815 °F &
102 psia/392 °)F
33.7
Siemens
V64.3A
34.8
Alstom
GTX100
P&WFT8+
36.5
37.2
GE
LM6000PD
40.9
598 °C
1108 °F
595 °C
1103 °F
554 °C
1029 °F
486 °C
907 °F
456 °C
853 °F
1.51
1.43
1.26
1.11
0.89
1.42
1.34
1.16
0.96
0.74
1.52
1.44
1.29
1.17
0.97
If a cogeneration plant needs to satisfy a steam load with an energy content Q, and a power load W, then the
ratio of Q/W can be used as a guideline for the selection of the configuration.
If Q/W > σ, supplementary firing would be needed to increase steam production, and a simple GT/HRB
configuration without a steam turbine is indicated. If, on the other hand, Q/W < σ, the GT exhaust energy
will not be fully utilised with a simple GT/HRB configuration, and a condensing steam turbine should be
included to take full advantage of the steam generation potential of the GT exhaust.
Since practical plants will typically operate with variable loads and conditions, a good design should never
be selected with parameters that place it on the edge of the range for which a configuration is suitable, as
further discussed below. Thus, a GT/HRB configuration is best if Q/W sufficiently exceeds σ, and a
configuration with a condensing steam turbine is best when Q/W is sufficiently below σ. The range of Q/W
for which each configuration is suitable is best understood by discussing the design considerations for each
cogeneration configuration.
18.1.2 DESIGN CONSIDERATIONS FOR GT/HRB COGENERATION PLANTS
An unfired HRB would produce Q/W ≈ σ. With a duct burner exit temperature in the 800 °C (1472 °F)
range, the upper bound for conventional HRB designs with insulated liners, the [Q/W] ratio is ≈ 2.5. Thus,
the design range for a GT/HRB configuration is σ < Q/W < 2.5.
As discussed above, the boundaries of the design range are not good design-points, since the operating
flexibility would be limited.
Suppose that a simple GT/HRB configuration without a steam turbine were used for a nominal Q/W ≈ σ. If
the operating steam demand increases, supplementary firing can be used to meet it, but if the steam demand
decreases, there is no satisfactory way to reduce the supply. A fixed HRB in a given hot exhaust stream will
dictate its own steam production, and the only way to reduce it is by curtailing the hot gas supply. This may
be done by modulating a diverter damper to spill some of the GT exhaust gas into a by-pass stack, a
cumbersome and inefficient procedure; or by part-loading the GT, undesirable or uneconomical if its power
output is needed. Another option is not to curtail the exhaust gas supply and to allow the HRB to generate
steam at its own rate, then to send the excess steam to a dump condenser, to recover the water. The
additional equipment needed by this option is less expensive than a by-pass stack and damper, since the
dump condenser can be at relatively high-pressure and operate at a relatively high temperature difference,
making it compact and more reminiscent of a feedwater heater than of a typical condenser.
18-2
 M. A. Elmasri, 1990-2002
Chapter 18: Cogeneration System Selection
Regardless of the measures used, it is costly and inefficient to get rid of unwanted steam generation. It is
therefore best to design a simple GT/HRB cogeneration plant with some supplementary firing at the nominal
design condition, to allow decreasing as well as increasing steam supply at off-design by modulating the
duct burner fuel flow in either direction.
Likewise, if a GT/HRB configuration were used for a nominal design where Q/W ≈ 2.5, it would correspond
to a level of supplementary firing at the upper bound of typical practice with a lined-duct boiler design. The
level of supplementary firing could be reduced in operation to meet a reduced steam demand, but could not
be increased to meet an increased steam demand.
Thus, the most desirable design range for a GT/HRB configuration without a steam turbine is
(σ + 0.3) < Q/W < 2.2
The upper bound for that range can be extended if specialised HRB designs are considered, such as with
ceramic liners or water-walls. With a duct burner exit temperature in the 1100 °C (2012 °F) range, which
would require a boiler with water-wall construction, the [Q/W] ratio can reach ≈ 3.5 for a modern GT of
either type. If the design calls for an even higher [Q/W] ratio, a plant based on gas turbines is unsuitable,
and a conventional boiler with a back-pressure steam turbine should be used.
18.1.3 DESIGN CONSIDERATIONS FOR COGENERATION PLANTS WITH A CONDENSING STEAM
TURBINE
A condensing combined cycle is the most efficient means of using a gas turbine to produce pure power,
without any cogeneration steam. It follows that if the amount of steam needed for process were relatively
small, the condensing combined cycle could be modified to extract the relatively small amount of
cogeneration steam from the steam turbine. Depending on the proximity of the process pressure to a value
suitable for one of the pressures of the heat recovery boiler, the steam may be bled directly from the HRB.
A condensing steam turbine should have enough steam to condense. For instance, if one were designed for a
nominal operating condition at which the condensing flow is just 10% of the inlet flow (with 90% extracted
for process), it may function perfectly at that design condition but is unlikely to function well at off-design
conditions. For instance, if the steam being extracted for process were reduced by half, the flow available
for the condensing section would increase by 550% (55% condensing vs. 10% condensing at nominal
conditions). This means that either (a) the condensing section would need to be grossly oversized relative to
its nominal capacity, in which case it would be expensive as well as inefficient at the nominal design
condition; or (b) at such an off-design condition the excess flow would be bled from the turbine and sent to a
dump condenser, because the condensing turbine section would be unable pass it. Problems of a similar
nature would occur if the process steam demand were to double.
The sizing problem of a condensing combined cycle can be ameliorated by using duct burners. If the
nominal design condition had the duct burners on, for instance, they could be turned down when process
steam demand dropped and vise versa. Such a solution can be effective, but begs the question as to whether
it wouldn't be cheaper and as efficient to forgo the condensing section of the turbine altogether, using a
back-pressure turbine instead.
A reasonable rule-of-thumb for selecting a condensing cogeneration combined cycle is that about half the
steam generation capacity should be condensed at nominal design conditions. This rule leads to the
guideline that the most desirable design range for cogeneration with a condensing/extraction steam turbine is
0 < Q/W < 0.6
18-3
 M. A. Elmasri, 1990-2002
Chapter 18: Cogeneration System Selection
18.1.4 DESIGN CONSIDERATIONS FOR COGENERATION PLANTS WITH A BACK-PRESSURE STEAM
TURBINE
For the intermediate range of Q/W, between the upper bound for a condensing cycle (Q/W ≈ 0.6) and the
lower bound for a GT/HRB configuration (Q/W ≈ σ), a back-pressure (non-condensing) steam turbine may
be used. The steam is generated at a higher pressure than needed for the process, then expanded through the
BPST to the process pressure, generating some additional power (and lowering the Q/W ratio).
Including a BPST makes more economic sense with lower process pressures and larger systems, since if its
power output is very small, the added complexity would be hard to justify. For a large plant (GT output >
150 MW say) where efficiency is at a premium, a back pressure turbine with a pressure ratio as low as 2:1
may be justified. Thus a back-pressure cycle may be reasonable with process pressures up to 750 psi (50
bar). For medium plants (GT output in the 60 MW range), one would normally use a back-pressure cycle if
one had at least a 4:1 available pressure ratio, i.e. if process steam pressure were below about 300 psi (20 bar
say). For small plants (GT output in the 10 MW range), one would normally use a back-pressure cycle only
if process steam pressure were below about 50 psi (4 bar).
18.2
PERFORMANCE OF VARIOUS COGENERATION ARRANGEMENTS
The various efficiency definitions for cogeneration plants were defined in the introduction. Below, we shall
present typical examples of the various arrangements, showing their characteristics and efficiencies.
Cogeneration with a Typical Heavy-Duty GT
Efficiency, %
20 bar (290 psia) process steam
95
90
85
80
75
70
65
60
55
50
45
40
35
30
25
CCC-U
CCC-F
1PBPCC
2PBPCC
GT/HRB
CCC-U
CCC-F
1PBPCC
2PBPCC
GT/HRB
0.0
0.5
1.0
1.5
2.0
2.5
Plant Net Heat to Power Ratio (Q/W)
Figure 1. Cogeneration Cycle Selection Map, heavy-duty GT and medium-pressure process steam
Figure 1 represents designs of cogeneration plants based on a typical heavy duty gas turbine. The lower
curves show electrical efficiency and the upper curves show total (CHP) efficiency. A GE Frame 6F is used
in these calculations, but since all results are presented as dimensionless ratios, they generally apply to any
gas turbine with an efficiency of about 34% (with losses) and en exhaust temperature of about 600 °C (1100
°F). Typical “medium pressure” (MP) cogeneration steam is assumed, at 20 bar/250 °C (290 psia/482 °F).
18-4
 M. A. Elmasri, 1990-2002
Chapter 18: Cogeneration System Selection
18.2.1 CONDENSING COMBINED CYCLES
At low Q/W, condensing combined cycles (CCC) with process steam extracted from the steam turbine are
used. The caption CCC-U refers to an unfired condensing combined cycle, and the caption CCC-F refers to
a cycle with design-point supplementary firing. The level of firing used in the latter designs is fixed at 680
°C (1256 °F), to provide greater off-design flexibility. The lines labelled CCC end at the Q/W point where
roughly half the generated steam is condensed. Figure 2 shows the heat balance diagram for the last CCC-U
point. Typical assumptions are used for all design-parameters, in conjunction with a conservative, twopressure, non-reheat cycle design.
GT PRO 10.3.5 Maher Elmasri
Net Power 90277 kW
LHV Heat Rate 8159 kJ/kWh
1.01 p
15 T
60 %RH
202 m
0 m elev.
206.1 m
1X GE 6101FA
14.85 p
374 T
14.11 p
1288 T
75.47 %N2+Ar
12.98 %O2
3.541 %CO2+SO2
8 %H2O
1.04 p
597 T
206.1 M
69143 kW
1p
15 T
202 m
23524 kW
CH4 4.088 m
LHV= 204612 kWth
25 T
62 T
37.31 M
95 T
60 p
2.428 M 540 T
30.4 M
71 T
V5
20 p 250 T 19.3 M
1.054 p
101 T
0.05 p
33 T
17.58 M
61.8 p 542 T
6 p 258 T
37.3 M
0.425 M
0.4293 M
4.127 M
33 T
2.428 M
LTE
LPB
IPE2
IPB
FW
HPE2
IPS1
HPE3
IPS2
HPB1
HPS3
595 T
206.1 M
107 T
206.1 M
1.054 p
95 T
37.31 M
1.208 p
105 T
0.425 M
6.614 p 6.614 p
160 T 163 T
37.3 M 4.127 M
131
135
178
p[bar], T[C], M[kg/s], Steam Properties: Thermoflow - STQUIK
64.92 p 6.487 p 63.96 p 6.36 p 63.96 p
228 T 277 T
238 T
260 T 280 T
30.71 M 4.127 M 30.71 M 4.127 M 30.4 M
217
264
267
293
295
61.8 p
542 T
30.4 M
501
595
Figure 2. Heat balance diagram for unfired CCC with process extraction at a rate which leaves about half the steam
generated to expand to the condenser.
The steam extracted does not contribute to steam turbine power between the extraction (20 bar) and the
condenser (0.05 bar) pressure, so ST output is reduced with extraction. The combined cycle design without
any process extraction at all, the leftmost point in Fig. 1, has a net electrical efficiency of 51.45% and a Q/W
of zero. The plant shown in Fig. 2 has a net electrical efficiency of 44.12% and a Q/W of 0.54. Both have
the same fuel input. Thus, the difference between the two designs is a sacrifice of electrical power output
equal to 51.45 - 44.12 = 7.33% of fuel input, in return for gaining a net useful heat output of 0.54*44.12 =
23.8% of fuel input. The thermal output gained is 3.25 times the electrical output lost. Thus, the increase in
total (CHP) efficiency is 2.25 times the decrease in electrical efficiency. CHP efficiency increases from
51.45% with zero Q/W, to 67.94% with Q/W=0.54, an increase of 16.49 points (2.25 times the drop of 7.33
points in electrical efficiency).
18-5
 M. A. Elmasri, 1990-2002
Chapter 18: Cogeneration System Selection
18.2.2 BACK-PRESSURE COMBINED CYCLES
For a higher Q/W ratio than suitable for condensing turbine cycles, one would use a back-pressure turbine.
Fig. 3 illustrates such a cycle, with all steam generation at an HP of 80 bar (1160 psia), which then passes
through the turbine and exhausts to the process pressure of 20 bar (290 psia). The conditions assumed for
this design are a “perfect match” between the steam generation capability of the unfired exhaust gases and
the process requirement, with all the steam that can be generated with a reasonable pinch, 20 °C (36 °F) ∗ in
the example, being sent to the process. Though dubbed a “perfect match”, such a design is awkward to
operate at off-design, since in the event of a reduction in process demand, there is no satisfactory way of
utilising the excess exhaust energy, whereas a design with supplementary firing at the nominal condition
would provide the flexibility of turning the duct burners down to reduce steam production.
Gross Power
Net Power
Aux. & Losses
Gross Heat Rate
Net Heat Rate
Gross Electric Eff.
Net Electric Eff.
Fuel LHV Input
Fuel HHV Input
Net Process Heat
77883 kW
76017 kW
1866 kW
9482 kJ/kWh
9714 kJ/kWh
37.97 %
37.06 %
205127 kWth
227612 kWth
87019 kWth
Ambient
1.013 P
15 T
60% RH
80 p
400 T
35.14 M
Stop Valve
8827 kW
21.4 p
244.5 T
34.79 M
IP Steam
34.79 M
1.013 p
170.7 T
206.1 M
HP
CH4
4.099 M
205127 kWth LHV
HPB
85.7 p
299.8 T
35.14 M
1.003 p
15 T
202 M
537.8 T
319.8 T
1.038 p
597.6 T
206.1 M
88.57 T
35.49 M
GE 6101FA
@100%load
69056 kW
Figure 3. Simplified heat balance diagram for unfired single-pressure BPCC, with all the steam generated with a
reasonable design pinch passing through the BPST then going to process, after desuperheating.
Fig. 4 shows the corresponding temperature profile. Had this plant been designed for a lower process
requirement, the pinch would have been greater than its minimum acceptable value (assumed at 15 °C or 27
°F) and the stack would have been hotter, wasting energy. Q/W for this example is 1.145. A lower Q/W
would waste heat up the stack at the nominal point. This design corresponds to the leftmost point of the
1PBPCC (single-pressure boiler, back-pressure combined cycle) family of design points, shown on Fig. 1.
Note that a 20 °C (36 °F) pinch is used in back-pressure cycles and GT/HRB cycles, since these should normally be designed with
supplementary firing, whereas a 15 °C (27 °F) pinch is used with condensing cycles, since these should normally be designed
without (appreciable) supplementary firing. When supplementary firing is turned off at reduced loads, the pinches are automatically
reduced, so it is generally economical to use larger pinches if the nominal sizing point includes significant supplementary firing.
∗
18-6
 M. A. Elmasri, 1990-2002
Chapter 18: Cogeneration System Selection
GT PRO 10.3.2 Maher Elmasri
700
Net Power 76017 kW
LHV Heat Rate 9714 kJ/kWh
HRSG Temperature Profile
HPE3
33030
HPS3
13801
HPB1
50420
0
600
Q kW
5
500
TEMPERATURE [C]
400
7
300
200
UA kW/C
100
578.5
64.95
704.6
0
0
10
20
30
40
50
60
70
80
90
100
HEAT TRANSFER FROM GAS [.001 X kW]
Figure 4. Heat recovery profile for the design of Fig. 3.
For the back-pressure system to produce more process steam, two options are possible. The first, obviously,
is to use supplementary firing, the second is to use dual-pressure heat recovery, since Fig. 4 shows that the
gases at the back of the HRB are still hot enough to produce some additional steam at process pressure. We
shall discuss supplementary firing first.
Fig. 5 shows the single-pressure cycle design with supplementary firing to 800 °C (1472 °F), the assumed
maximum for common HRB constructions. At that level of firing, the heat recovery is no longer constrained
by the HP pinch, but by the stack, assumed to be 100 °C (212 °F) in the calculation. Fig. 6 shows the
corresponding HRB profile, with the pinch being well above the minimum permissible value. The Q/W
ratio is about 1.795. This design corresponds to the rightmost point of the 1PBPCC (single-pressure boiler,
back-pressure combined cycle) family of design points, shown on Fig. 1.
The best designs for single-pressure BPCC’s are those with a level of supplementary firing between the
designs of Fig. 3 and Fig. 5, say with firing to about 700 °C (1292 °F) at the nominal conditions, for a Q/W
of about 1.5. Such a design allows modulation of steam production via the duct burners, with the swing of
±100 °C (±180 °F) in exhaust temperature after the burner enabling a swing of roughly ±20% in steam
production.
The line on Fig. 1 representing the 1PBPCC family of design points changes slope between its leftmost point
(the unfired design of Fig. 3) and its rightmost point (the fully-fired design of Fig. 5). This change of slope
is at the level of firing at which the minimum stack temperature is attained. To the left of this point, designs
with increasing supplementary firing benefit from the reduction in stack temperature, since heat recovery is
constrained by the HP pinch. To the right of this point, designs with increasing duct burner exit
temperatures do not get this additional benefit, thus the slope of the efficiency curves is reduced.
18-7
 M. A. Elmasri, 1990-2002
Chapter 18: Cogeneration System Selection
Gross Power
Net Power
Aux. & Losses
Gross Heat Rate
Net Heat Rate
Gross Electric Eff.
Net Electric Eff.
Fuel LHV Input
Fuel HHV Input
Net Process Heat
84687 kW
82175 kW
2511.7 kW
11001 kJ/kWh
11337 kJ/kWh
32.73 %
31.75 %
258781 kWth
287147 kWth
147479 kWth
Ambient
1.013 P
15 T
60% RH
80 p
400 T
59.8 M
Stop Valve
15631 kW
21.4 p
240.8 T
59.2 M
IP Steam
59.2 M
1.013 p
99.98 T
207.2 M
HP
CH4
4.099 M
205127 kWth LHV
HPB
85.7 p
299.8 T
59.8 M
1.003 p
15 T
202 M
707 T
350.5 T
1.038 p
597.6 T
206.1 M
88.57 T
60.39 M
GE 6101FA
@100%load
69056 kW
CH4
1.072 M
53654 kWth LHV
Figure 5. Simplified heat balance diagram for single-pressure BPCC, with supplementary firing to 800 °C (1472 °F).
Net Power 82175 kW
LHV Heat Rate 11337 kJ/kWh
GT PRO 10.3.2 Maher Elmasri
HRSG Temperature Profile
HPE3
56208
900
HPS3
23485
HPB1
85800
0
800
Q kW
5
700
TEMPERATURE [C]
600
500
400
7
300
200
UA kW/C
100
506.4
58.98
2179.5
0
0
20
40
60
80
100
120
140
160
HEAT TRANSFER FROM GAS [.001 X kW]
Figure 6. Heat recovery profile for the design of Fig. 5.
18-8
180
200
 M. A. Elmasri, 1990-2002
Chapter 18: Cogeneration System Selection
Returning to the unfired design of Fig. 3. The second path of obtaining additional process flow will now be
discussed, viz the use of dual-pressure heat recovery. Adding an IP drum recovers more heat from the gases
before the stack, lowering the stack temperature. The IP steam is generated at the process pressure (plus
losses) and supplements the process steam produced by the HP, after it passes through the BPST. Thus
additional process steam can be produced without supplementary firing, resulting in greater efficiency.
Figure 7 illustrates the maximum unfired process steam production possible with a dual-pressure system, for
the assumed pinch points of 20 °C (36 °F). Its temperature profile is shown in Fig. 8. Its Q/W ratio is 1.22.
This design corresponds to the leftmost point of the 2PBPCC (two-pressure boiler, back-pressure combined
cycle) family of design points, shown on Fig. 1.
If one needed more steam from a dual-pressure design, supplementary firing may be used. As shown in
Chapter 10, supplementary fired dual-pressure designs can produce more steam at the HP relative to the IP,
making the IP contribution to heat recovery less significant. Indeed, once the level of firing reaches the
transition temperature (see Chapter 10), the dual-pressure system is superfluous. Thus, the 2PBPCC family
of design points merges into the 1PBPCC family on Fig. 1, as Q/W increases with supplementary firing.
Gross Power
Net Power
Aux. & Losses
Gross Heat Rate
Net Heat Rate
Gross Electric Eff.
Net Electric Eff.
Fuel LHV Input
Fuel HHV Input
Net Process Heat
77883 kW
75862 kW
2020.5 kW
9482 kJ/kWh
9734 kJ/kWh
37.97 %
36.98 %
205127 kWth
227612 kWth
93182 kWth
Ambient
1.013 P
15 T
60% RH
80 p
400 T
35.14 M
Stop Valve
8827 kW
21.4 p
244.5 T
34.79 M
IP Steam
37.34 M
1.013 p
142.8 T
206.1 M
IP
HP
IPB
21.4 p
215.8 T
2.548 M
HPB
85.7 p
299.8 T
35.14 M
257.5 T
235.8 T
537.8 T
319.8 T
CH4
4.099 M
205127 kWth LHV
1.003 p
15 T
202 M
1.038 p
597.6 T
206.1 M
88.62 T
38.06 M
GE 6101FA
@100%load
69056 kW
Figure 7. Simplified heat balance diagram for unfired dual-pressure BPCC, with all the HP & IP steam generated with
reasonable design pinches going to process. The HP steam first expands through the BPST, before joining the IP.
18-9
 M. A. Elmasri, 1990-2002
Chapter 18: Cogeneration System Selection
GT PRO 10.3.2 Maher Elmasri
700
IPB
4809
IPE2
20365
Net Power 75862 kW
LHV Heat Rate 9734 kJ/kWh
HRSG Temperature Profile
HPE3
13931
HPS3
13801
HPB1
50420
0
600
Q kW
5
500
TEMPERATURE [C]
400
6
300
9
11
200
UA kW/C
100
569.9
164.5
0
0
20
578.5
64.95
427.7
40
60
80
100
120
HEAT TRANSFER FROM GAS [.001 X kW]
Figure 8. Heat recovery profile for the design of Fig. 7.
18.2.3 GT/HRB CYCLE
The family of designs designated “GT/HRB” on Fig. 1 represent cogeneration cycles without a steam
turbine, where process steam is generated directly at the desired pressure. The condition of maximum
unfired steam production with a reasonable pinch is the leftmost point in that family, with a Q/W ratio of
1.54. Supplementary firing can generate more steam, and the rightmost point in that family, with
supplementary firing to 800 °C (1472 °F) has a Q/W ratio of 2.44.
With supplementary firing to 1000 °C (1832 °F), which would require ceramic HRB liners (or water-walls),
a Q/W ratio of about 3.3 can be obtained, this is considered the approximate upper bound of practical GTbased cogeneration. Higher levels of firing with HRB water-walls end at the physical limit of exhaust
oxygen depletion, reached at a Q/W of about 5, but such cogeneration applications may be better served with
conventional boilers and back-pressure steam turbines, without gas turbines.
18.2.4 DEADBANDS BETWEEN CONFIGURATIONS
Figure 1 shows a “deadband” for Q/W between roughly 0.6 and 1.2. In this range, a condensing combined
cycle would have too little flow to the condenser, making it awkward or inefficient to design and operate
with variable steam demand. A non-condensing cycle would not take full advantage of the GT exhaust heat.
Although one could certainly design a condensing combined cycle in this region, with various measures to
enhance operating flexibility, such as heavy supplementary firing and steam by-pass around the turbine to
the process when needed, or using an oversized condensing section with auto-extraction valves, it is most
economical and efficient to avoid this deadband, if possible.
Since most cogeneration plants have alternate sources of power (such as buying from the grid), alternate
sinks of power (such as selling to the grid), and alternate sources of steam (such as package boilers), it
should be fairly easy to avoid designs in this deadband. If the ideal plant had a nominal Q/W within this
18-10
 M. A. Elmasri, 1990-2002
Chapter 18: Cogeneration System Selection
band, and the Q was the more important constraint, one may consider building a larger facility (more W) to
reduce the Q/W to below 0.5, the comfort range for condensing plants. In this case one would need to sell
surplus power. Alternatively, one could build a smaller plant (less W) to increase Q/W to above 1.2, the
comfort range for a back-pressure cycle. In this case, one would need to purchase power to make up for the
shortfall.
18.2.5 OVERLAP BETWEEN CONFIGURATIONS
Figure 1 shows an overlap between three alternate configurations in the region of Q/W around 1.6. Designs
in this region can be single-pressure BPCC, dual-pressure BPCC, or GT/HRB. All three have comparable
efficiencies, with the BPCC’s having a slight advantage in efficiency, which may be offset by their higher
cost and complexity. In this region, all three should be evaluated to find the economic optimum.
The availability of a gas turbine (or a multiplicity of identical gas turbines) with a power output best suited
to the Q and W demand is another important factor. Figure 9 illustrates the GT output as a percentage of the
plant net output for all the cycle designs given in Fig. 1.
Cogeneration with a Typical Heavy-Duty GT
20 bar (290 psia) process steam
GT Gross Output as % of Plant Net
105
100
95
90
CCC-U
85
CCC-F
80
2PBPCC
1PBPCC
GT/HRB
75
70
65
60
0.0
0.5
1.0
1.5
2.0
2.5
Plant Net Heat to Power Ratio (Q/W)
Figure 9. Gas Turbine output as a percentage of plant net output for various cycle designs (MP cogen steam)
Suppose, for example, the proposed plant design needs to supply a nominal W of 50 MW (of electrical
energy) and a nominal Q of 80 MW (of thermal energy), giving a Q/W of 1.6. If a GT/HRB system were
chosen, Fig. 9 shows that the GT nominal gross output would need to be 102% of the plant’s, i.e. 51 MW. If
a BPCC were chosen, Fig. 9 shows that the GT nominal gross output would need to be 86% of the plant’s,
i.e. 43 MW. The availability of a desirable GT model of either 51 MW or 43 MW would influence the
decision on which configuration to choose.
18.2.6 EFFECT OF COGENERATION STEAM PRESSURE
Figures 10 & 11 show the cycle selection map and the GT contribution to the net plant output for each
design, with typical “low pressure” (LP) process steam, at 3 bar/175 °C (44 psia/347 °F). Because the backpressure cycles produce much more power with a low BPST exhaust pressure, another deadband appears
18-11
 M. A. Elmasri, 1990-2002
Chapter 18: Cogeneration System Selection
between the fully-fired BPCC and the unfired GT/HRB designs. Also, the combined cycles, whether
condensing or back-pressure are more efficient, but the GT/HRB cycle is not.
Cogeneration with a Typical Heavy-Duty GT
Efficiency, %
3 bar (44 psia) process steam
95
90
85
80
75
70
65
60
55
50
45
40
35
30
25
CCC-U
CCC-F
1PBPCC
2PBPCC
GT/HRB
CCC-U
CCC-F
1PBPCC
2PBPCC
GT/HRB
0.0
0.5
1.0
1.5
2.0
2.5
Plant Net Heat to Power Ratio (Q/W)
Figure 10. Cycle Selection Map similar to that of Fig. 1, but with low process steam pressure
Cogeneration with a Typical Heavy-Duty GT
3 bar (44 psia) process steam
GT Gross Output as % of Plant Net
105
100
95
90
CCC-U
85
CCC-F
80
2PBPCC
1PBPCC
GT/HRB
75
70
65
60
0.0
0.5
1.0
1.5
2.0
2.5
Plant Net Heat to Power Ratio (Q/W)
Figure 11. Gas Turbine output as a percentage of plant net output for various cycle designs (LP cogen steam)
18-12
 M. A. Elmasri, 1990-2002
Chapter 18: Cogeneration System Selection
18.2.7 COGENERATION WITH A TYPICAL AERODERIVATIVE
Cogeneration with a Typical Aeroderivative
20 bar (290 psia) process steam
95
90
85
CCC-U
80
CCC-F
Efficiency, %
75
1PBPCC
70
2PBPCC
65
GT/HRB
60
CCC-U
55
CCC-F
50
1PBPCC
45
2PBPCC
40
GT/HRB
35
30
25
0.0
0.5
1.0
1.5
2.0
2.5
Plant Net Heat to Power Ratio (Q/W)
Figure 12. Cycle Selection Map similar to that of Fig. 1, but with Aeroderivative GT
Cogeneration with a Typical Aeroderivative
20 bar (290 psia) process steam
GT Gross Output as % of Plant Net
105
100
95
90
CCC-U
85
CCC-F
80
2PBPCC
1PBPCC
GT/HRB
75
70
65
60
0.0
0.5
1.0
1.5
2.0
2.5
Plant Net Heat to Power Ratio (Q/W)
Figure 13. Aeroderivative GT output as a percentage of plant net output for various cycle designs (MP cogen steam)
18-13
 M. A. Elmasri, 1990-2002
Chapter 18: Cogeneration System Selection
Figures 12 and 13 show the cycle selection maps and relative GT contributions in each for a typical
aeroderivative GT, with medium pressure cogeneration steam, at 20 bar/250 °C (290 psia/482 °F). Figures
14 and 15 show the same, but with low pressure cogeneration steam, at 3 bar/175 °C (44 psia/347 °F). A GE
LM6000 PD is used in these calculations, but since all results are presented as dimensionless ratios, they
generally apply to any gas turbine with an efficiency of about 41% (with losses) and en exhaust temperature
of about 450 °C (840 °F).
The range of possible cycle designs with supplementary firing is much wider than with a heavy duty GT.
For instance, the BPCC range extends down to a Q/W of about 0.6, touching the range for the CCC’s. The
range for the GT/HRB cycles extends down to a Q/W of about one, widely overlapping the range for the
BPCC’s. The design deadbands of Figs. 1 and 10 disappear. The wider design range is created by the wider
range between the aeroderivative’s low exhaust temperature, of about 450 °C (842 °F) and the assumed
“fully-fired” duct burner exit temperature, of 800 °C (1472 °F). This range, at about 350 °C (630 °F), is
75% greater than the 200 °C (360 °F) range between the unfired and “fully-fired” exhaust temperature with a
heavy duty GT, creating a wider range of design options with a wider range of post-fired exhaust energy.
With supplementary firing to a given duct burner exit temperature, modern aeroderivative and heavy duty
machines will produce roughly the same amount of steam per kW of GT power, but the aeroderivative will
require more duct burner fuel to attain the same final exhaust temperature. This is because with
supplementary firing to the same duct burner exit temperature, the amount of steam that can be generated is
proportional to the exhaust mass flow. Thus the amount of steam generated per kW of gas turbine power
would be inversely proportional to the specific power of the gas turbine. Specific Power (kW per unit
airflow) and thus exhaust mass flow per kW of GT output, is comparable for modern heavy duty and
aeroderivative machines (about 350 kW per kg/s or 160 kW per lb/s). When supplementary firing is used
with an aeroderivative to increase Q/W to be in the same range as with a heavy duty GT, plant efficiencies
with either GT type become similar.
Cogeneration with a Typical Aeroderivative
3 bar (44 psia) process steam
95
90
85
CCC-U
80
CCC-F
Efficiency, %
75
1PBPCC
70
2PBPCC
65
GT/HRB
60
CCC-U
55
CCC-F
50
1PBPCC
45
2PBPCC
40
GT/HRB
35
30
25
0.0
0.5
1.0
1.5
2.0
2.5
Plant Net Heat to Power Ratio (Q/W)
Figure 14. Cycle Selection Map similar to that of Fig. 12, but with low process steam pressure
18-14
 M. A. Elmasri, 1990-2002
Chapter 18: Cogeneration System Selection
Cogeneration with a Typical Aeroderivative
3 bar (44 psia) process steam
GT Gross Output as % of Plant Net
105
100
95
90
CCC-U
85
CCC-F
80
2PBPCC
1PBPCC
GT/HRB
75
70
65
60
0.0
0.5
1.0
1.5
2.0
2.5
Plant Net Heat to Power Ratio (Q/W)
Figure 15. Aeroderivative GT output as a percentage of plant net output for various cycle designs (LP cogen steam)
The wider range of post-firing exhaust temperatures available with an aeroderivative enhances operating
flexibility, but complicates the design. For instance, with a 450 °C (842 °F) aeroderivative exhaust and a
maximum post-firing temperature of 800 °C (1472 °F), a plant designed with its nominal condition at the
midpoint of this range, i.e. fired to 625 °C (1157 °F), can modulate steam production via the duct burners,
with the swing of ±175 °C (±315 °F) in post-combustion temperature, enabling a swing of roughly ±45% in
steam production. This compares with a range of ±20% for the typical heavy duty GT. The gain in offdesign operational flexibility, however, is at the cost of a much more complex design. Sizing a boiler to
operate flexibly and efficiently across a wide range of post firing temperatures is difficult, as discussed in
Chapter 9, §6.c. Sizing a steam turbine to operate across a wide range of throttle steam flows is likewise
complicated, as discussed below.
18.3
DESIGNING
RATES
A
STEAM TURBINE
TO
OPERATE ACROSS
A
RANGE
OF
FLOW
The simplest construction of a steam turbine is one without active pressure controls, i.e. one that will operate
with sliding pressure. For a such a steam turbine, the flow-passing capacity of each set of nozzles sets the
proportionality constant between steam mass flow and pressure. If more flow were “pushed” through the
nozzles, the pressure would rise proportionately, and vise versa. Turbines operated in sliding pressure,
without active pressure controls, are also most efficient, since the proportionality between pressure and flow
keeps volumetric flow rate (nearly) constant, hence velocity triangles remain undistorted from their designpoint shapes. Furthermore, throttling losses created by pressure control valves are avoided.
Unfortunately, it is very difficult to use such a simple and efficient steam turbine with wide variations in
steam flow rates. For example, assume a cycle were designed with an optimum throttle pressure of 80 bars,
say, at the nominal 100% steam flow rate, and the steam turbine inlet nozzles sized on this basis. If the plan
18-15
 M. A. Elmasri, 1990-2002
Chapter 18: Cogeneration System Selection
was to use the duct burner at off-design to modulate steam flow by ±45%, as discussed above, the ST inlet
pressure would also swing by roughly ±45%. At full firing, ST inlet pressure would climb to ≅ 1.45 x 80 =
116 bars, requiring all components to be sized for this pressure, which would increase the cost considerably.
At no firing the ST inlet pressure would drop to ≅ 0.55 x 80 = 44 bars, causing higher velocity in pipes and
superheaters, as well as steaming in the HRB’s economisers.
An alternative design would be to size the turbine for the maximum flow at the nominal pressure. If the ST
nozzles were sized to pass 145% of nominal flow at the nominal pressure of 80 bars, the pressure would
slide down to ≅ 100/145 x 80 = 55 bars at the nominal flow, if left uncontrolled. To control it to 80 bars at
the nominal flow, a valve upstream of the ST inlet would have to be partially closed, throttling the steam
from 80 bars upstream of the valve to 55 bars downstream of it, in the ST nozzle “bowl”. The efficiency
loss from this throttling would degrade performance at the nominal flow. This degradation can be
minimised, but not eliminated, by using multiple valves or variable area nozzles, but these are expensive
constructions.
The same difficulties are encountered at steam turbine extraction ports. If the flow entering the stages
downstream of the extraction ports were reduced, the pressure would fall. The reduction in “flow-tofollowing-stage” can be either due to a reduction in upstream supply, or due to an increase in extracted flow.
To maintain the pressure at the extraction port, a control device may be installed to constrict the flow to the
following stages, i.e. an auto-extraction steam turbine design. The control device may be a simple throttle or
grid valve, or a more elaborate multi-valve system or variable area nozzles, as discussed in Chapter 12. Any
method of controlling the extraction pressure will reduce efficiency and increase cost. An alternative
approach is to equip the turbine with multiple extraction ports, all of the simple, sliding-pressure type. If the
pressure at an extraction port slides too low, an automatic control would shut a valve on the pipe taking the
steam from this port, and open a valve on the pipe from an upstream extraction port, with an adequate
pressure. Although this method is simpler and less expensive, it also has the efficiency penalty of always
sourcing the steam from a port with a surplus pressure, since the number of ports cannot be too large. In
most practical designs of this system, two sliding ports would be used, with a let-down from the ST inlet as a
third steam source of last resort.
The method of control used and the location of the extraction port(s) are all parameters which should be
optimised with a detailed study for each particular plant, since the optimum solution will depend on the load
duration curve, i.e. on the variation of steam demand with time over the plant’s duty cycle.
18-16
 M. A. Elmasri, 1990-2007
Chapter 19: Repowering
REPOWERING STEAM PLANTS WITH GAS TURBINES
Revised September, 2005, updated October, 2007
© Maher Elmasri 1999-2007
19.1
INTRODUCTION
Adding gas turbines to an existing steam plant can increase its efficiency or capacity or both. The gas
turbine exhaust energy needs to be effectively utilised to decrease the fuel consumption of the steam
boiler, or increase the power-producing steam in the steam turbine, or both.
There are three basic categories of repowering:
(a) The existing steam boiler is preserved and continues to be the sole source of steam to the original
steam turbine.
(b) The existing boiler is preserved and augmented by steam from a new gas turbine heat recovery boiler.
In this category, both the new and old sources of steam operate in parallel to power the original steam
turbine.
(c) The existing steam boiler is decommissioned and totally substituted by a new heat recovery boiler to
supply the original steam turbine.
If the existing boiler burns low-grade fuel (such as coal or residual oil), which cannot be utilised directly
in gas turbines, and if the plant needs to continue burning that low-grade fuel, then categories (a) and (b)
are the only choices. Otherwise, all three categories are worthy of consideration. Factors such as the
desired increase of capacity, the change in efficiency, and the availability of land at the site are then used
to decide on an approach.
In considering the repowering of a plant, one needs to ascertain that the main equipment being preserved
has a remaining service life commensurate with the new equipment being introduced, otherwise the
combination may not be economically sensible.
19.1.1 METHODOLOGY & MAIN ASSUMPTIONS
In what follows a comparison of the main performance characteristics of the different repowering
schemes is presented. The numerical analysis is based on a typical subcritical steam plant, with a single
reheat, seven feedwater heaters, and nominal steam conditions as follows:
Table 1. Assumed Nominal Steam Conditions of Existing Steam Plant
HPT inlet
Reheat steam
Crossover
Condenser
140 bar/545 ºC
30 bar/545 ºC
6 bar
.05 bar
2035 psia/1015 ºF
435 psia/1015 ºF
90 psia
1.5 "Hg
In an attempt to be as general as possible, the analysis is not size-specific or manufacturer-specific.
Rather than absolute size, the governing factor is the ratio of gas turbine size to original steam power
plant size, both expressed in MW. The descriptors of the relationship between the gas turbine and the
steam plant is therefore GT exhaust mass flow rate per MW (inverse of specific power), and GT exhaust
temperature. An additional, factor in overall repowered plant performance is the GT efficiency (or heat
rate). Thus, without being size-specific, the relevant GT parameters are its specific power, efficiency, and
exhaust temperature. The repowering results shown are therefore based on hypothetical, modern gas
turbines, one representing the high-pressure-ratio, high-efficiency, aeroderivative genre and one
representing the heavy-duty genre. The relevant characteristics assumed for each genre are as follows:
19-1
 M. A. Elmasri, 1990-2007
Chapter 19: Repowering
Table 2. Assumed Characteristics of Aeroderivative GT Genre
Specific Power
Heat Rate
Exhaust temperature
335 kW/kg/s
8888 kJ/kWh
460 ºC
152 kW/lb/s
8425 BTU/kWh
860 ºF
Table 3. Assumed Characteristics of Heavy Duty GT Genre
Specific Power
Heat Rate
Exhaust temperature
360 kW/kg/s
9960 kJ/kWh
575 ºC
163 kW/lb/s
9441 BTU/kWh
1067 ºF
The characteristics assumed for the aeroderivative genre are roughly similar to performance of GE
LM6000 or RR Trent. Those for the heavy-duty genre are consistent with "F-Class" machines from GE
and "A" Class machines from Siemens.
The specific power cited is based upon GT exhaust gas, rather than inlet air, flow rates. The exact size of
the GT is not assumed for any particular model, rather, the GT aggregate output is expressed as a
percentage of the steam plant original net output. Thus, the analysis can be applied to a range of steam
plant and gas turbine sizes.
Although the calculations are for assumptions typical of a seven-heater steam plant in the 350 MW class,
the results should be numerically applicable for most single-reheat steam units with 5-8 heaters and
between 80 and 600 MW. The results should still be applicable, but only as crude guidelines, for nonreheat, straight condensing plants and double-reheat, supercritical plants.
19.1.2 REPOWERING DEFINITIONS
In analysing the results, the following definitions were introduced for convenience:
Repowering Efficiency, η repowering , is the ratio of increased net power output to increased fuel heat
input:
η repowering =
kWnew − kWold
Qin , new − Qin ,old
.................................... (1)
Gas Turbine Leverage, λGT ,is the ratio of increased net power output to GT power output:
λGT =
where
kWnew − kWold
kWGT
........................... (2)
kWnew
kWold
Qin , new
Net plant output after repowering
Qin ,old
Total original steam plant fuel heat input
kWGT
Nominal gross output of gas turbine(s) added for repowering
Net original steam plant output
Total fuel heat input after repowering
Repowering Efficiency can be used to study the economics of repowering, since it is the effective
efficiency of the incremental capacity added by repowering. It should be analyzed as a trade-off with the
capital cost of repowering, which is the cost of obtaining that new increment of capacity. Naturally, it is
only meaningful in a direct way when the same fuel is used before and after repowering, such as when the
19-2
 M. A. Elmasri, 1990-2007
Chapter 19: Repowering
original steam plant was burning gas or light oil in its boiler. If the original steam plant was burning an
inexpensive low-grade fuel, and the additional gas turbine capacity will burn expensive, high-grade fuel,
then this thermodynamic definition must be used with caution, since it does not directly imply the
economic efficiency.
GT Leverage provides some insight into the costliness of the repowering project compared to that of a
new power plant. For instance, a stand-alone gas turbine has a leverage of 1 whereas a gas turbine
coupled with a new purpose-built combined cycle has a leverage of about 1.5. In repowering, when the
GT is integrated with an existing steam plant, it may result in an increase in the output of the original
steam turbine, in which case its leverage is greater than 1. Frequently, however, its integration
necessitates a derating of the original steam turbine, in which case its leverage is less than 1.
Generally speaking, it will be shown that repowering schemes which offer a high Repowering Efficiency
are associated with a low GT Leverage, and therefore high cost per incremental kW. Thus, one may find
schemes of adding incremental capacity at a very high incremental efficiency. Alas, such cases are
usually balanced by the fact that the capacity gain is much less than the output of the GT purchased. In
extreme cases, Repowering Efficiency can exceed 100% and GT Leverage can be negative !
19.2
REPOWERING WITH THE EXISTING STEAM BOILER
19.2.1 FEEDWATER REPOWERING
In this scheme, illustrated by Figure 1a, a gas turbine is installed to produce additional power. Its exhaust
energy is utilised to heat a portion of the feedwater of the original steam plant, which bypasses the normal
feedwater heaters. Thus, the feedwater heaters draw less steam from the turbine bleeds than in the
original steam plant, saving some steam which can expand to the condenser and produce additional steam
turbine power. The increased power output of the steam turbine augments the additional power added by
installing the new gas turbine.
The ability of the steam turbine to accommodate the additional steam flow at its low pressure end and its
ability to generate the additional power are both important to obtain maximum benefit from feedwater
repowering. If either the steam flow through the low pressure section or the power output reach their
limiting values, then the original boiler has to be run at part load. While this preserves the efficiency gain
obtained by repowering, it reduces the capacity gain.
The heat recovery “boiler” illustrated in Figure 1a consists of two water heating coils, and its
thermodynamic design is constrained by a pinch at the steam cycle’s deaerator (the contact heater,
designated “C” in Figure 1a). The gas turbine exhaust leaving the HP coil “FWA” has to be at a
temperature higher than the feedwater entering the coil by a reasonable value, 10-15 °C (18-27 °F) say.
The gas turbine exhaust leaving the LP coil “FWB” has to be at or above a minimum stack temperature,
suitable for the GT fuel. With a typical aeroderivative, the GT exhaust is at about 460 °C (860 °F) and
the deaerator of a conventional steam cycle is commonly at 120-160 °C (248-320 °F). Thus, most of the
GT exhaust gas energy is available above the pinch, to heat HP water in “FWA”, and a relatively small
portion is left over to heat LP water in “FWB”. For example, assuming a gas turbine exhaust of 460 °C, a
deaerator at 140 °C (after the feedpump temperature rise), and a 15 °C pinch, gives an available exhaust
temperature drop of 460-(140+15) = 305 °C for “FWA”. Assuming a 100 °C stack gives an available
exhaust temperature drop of (140+15)-100 = 55 °C for “FWB”. Thus, of the total exhaust heat recovery,
roughly 85% is available for heating the HP water in “FWA”, and 15% is left over to heat LP feedwater
in “FWB”.
19-3
 M. A. Elmasri, 1990-2007
Chapter 19: Repowering
Plant net power
Plant net HR (LHV)
Plant net eff (LHV)
Aux. & losses
Fuel heat input (LHV)
427645
8503
42.34
13463
1010084
kW
kJ/kWh
%
kW
kWth
Ambient
1.013 p
15 T
60% RH
6.253 p
357555 kW
HPT
G
4 LPTs
IPT
3600 RPM
140 p
547.6 T
263 M
29.8 p
528.3 T
255.5 M
0.05355 p
34.12 T
210.8 M
Stack
127.3 T
324.4 M
16.15 M Fuel
308.3 M Air
239 T
D
D
C
D
F
79%
2xGE LM6000PD
83553 kW
D
P
16.6%
Stack
124.2 T
452.2 T
246.7 M
G
FWA
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FWB
p [bar] T [C] M [kg/s]
Figure 1a. An illustration of Feedwater Repowering
Heating the feedwater in a conventional steam cycle plant by using gas turbine exhaust, whether as an
original design concept, or as a repowering modification, has been practised for many years, although not
at a great many plants. One of the first gas turbines ever installed for power generation in the USA was a
3.5 MW GE Frame 3, installed in 1949 at Oklahoma Gas & Electric’s Belle Isle plant, where its exhaust
heated the feedwater for a 35 MW steam turbine plant. This gas turbine is now on display at GE’s plant
in Schenectady, NY, and has been designated a historic engineering landmark by the ASME. Similar
installations were built in 1952 at Belle Isle (Frame 3) and 1961 at Liberal, KS (Frame 5).
More recently, the Danish company Energi E2 (formerly SK Power) has designed and built two plants
using “hybrid cycles” in which gas turbines heat feedwater for conventional steam cycles burning coal
and biomass. One of these is at the “Map Ta Phut” plant, in Thailand, completed in 1997, in which two
GE Frame 6B gas turbines are used per steam unit to heat the feedwater as well as reheat the steam. The
other is the “Avedore 2” plant in Denmark, completed in 2001, in which two Rolls Royce Trent gas
turbines are used to heat the feedwater for the supercritical, single-reheat, steam cycle. The Avedore 2
plant is designed to operate in conventional mode only, with extraction feedwater heating, with the gas
turbines off; or in “hybrid mode” with either or both gas turbines on, heating the feedwater and bypassing the extraction feedwater heaters.
19-4
 M. A. Elmasri, 1990-2007
Chapter 19: Repowering
Another recent example is in Belgium,
at Electrabel’s Langerlo plant, in which
two GE LM6000 gas turbines were
installed in feedwater repowering duty.
Figure 1b shows how compact the heat
recovery unit is, given that it consists
entirely of water-heating coils, without
the evaporators or superheaters that
would normally be installed in a
conventional HRB.
Figure 1b. An application of feedwater
repowering. The LM 6000 gas turbine’s
heat recovery unit at Langerlo (courtesy of
GE Energy)
Figure 2 shows the relative values of
key plant parameters as a function of
the size of the gas turbine(s) selected.
This figure is produced with gas turbine characteristics of a typical aeroderivative machine. Gas turbine
capacity of about 16% of the original plant will result in an increase of ST condensing flow to about
110% of nominal full load value; with ST power output increased to about 104% of nominal value. The
GT Leverage would be 122%, i.e. for each MW of gas turbine capacity installed, the plant's output would
increase by 1.22 MW. The total capacity increase for the plant would be about 20% and overall plant
heat rate improves by about 4%. Repowering efficiency, defined as the ratio of incremental power to
incremental fuel is about 53%.
Fig. 2 shows that aeroderivatives with a capacity of about 30% of the original plant would be able to heat
100% of the HP water, saving all the HP bleed steam. Using a greater gas turbine capacity would be at a
declining efficiency, since the additional GT exhaust energy would be used only for heating LP water,
thereby saving only LP bleed steam. One should note, however, that with the boiler at full rated load,
using a gas turbine capacity as large as 30% of the original plant is only possible if the ST can
accommodate 120% of the original steam flow to the condenser, and can generate 108% of its original
power output. A gas turbine capacity of 20% of the original steam plant’s is therefore more likely to be
the limit, since it would only increase ST exhaust flow by about 12% and ST generator output by about
5%.
The higher ratio of exhaust energy to power output of a heavy-duty gas turbine, relative to an
aeroderivative, allows a smaller plant capacity increase for a given, limiting LPT steam flow. The
efficiency improvement is also lower than for aeroderivatives. Thus, aeroderivatives are a better
19-5
 M. A. Elmasri, 1990-2007
Chapter 19: Repowering
thermodynamic match for feedwater repowering than heavy duty gas turbines.
Also, since
aeroderivatives are commonly available in the range 25-50 MW, a single gas turbine is well suited for
feedwater repowering of a steam unit in the range 125-300 MW.
140
130
120
110
100
Plant net kW
ST gross kW
%
90
ST exhaust flow
80
Plant net eff.
70
Repowering eff.
GT Leverage
60
HP FW bypass
LP FW bypass
50
40
30
20
10
0
5
10
15
25
20
30
35
40
GT Rating as % of Original Plant Net
Figure 2. Typical Feedwater Repowering results with aeroderivatives
Feedwater repowering is an attractive way of adding power to an existing steam plant at high efficiency
and moderate cost per incremental kW. The alterations to the original steam plant are minimal, simply
consisting of piping water around the heaters to the new heat recovery "boiler". If space permits, the new
gas turbine and heat recovery "boiler" can be erected close to the existing turbine hall, but if space
limitations require piping the water for several hundred meters, the system is still feasible. Necessary
downtime is very short or even non-existent if heater by-pass pipes and valves are already in place. This
type of repowering can add an incremental 50 MW at 52% efficiency to an existing 250 MW steam unit
at a cost of around $700/kW. A new 50 MW stand-alone combined cycle would achieve about the same
efficiency but would cost around $1100/kW.
19-6
 M. A. Elmasri, 1990-2007
Chapter 19: Repowering
19.2.2 DIRECT HOT WINDBOX REPOWERING
Plant net power
Plant net HR (LHV)
Plant net eff (LHV)
Aux. & losses
Fuel heat input (LHV)
400632
7963
45.21
11090
886150
kW
kJ/kWh
%
kW
kWth
Ambient
1.013 p
15 T
60% RH
5.212 p
310463 kW
HPT
4 LPTs
IPT
G
3600 RPM
B
140 p
542.5 T
227.3 M
A
25.27 p
550 T
215.9 M
0.04555 p
31.24 T
176.7 M
Stack
131.4 T
363.5 M
11.6 M Fuel
351.9 M 540 T Gas
235.4 T
D
A
1xKWU V84.2
101259 kW
D
D
D
C
A
34.2%
B
F
P
B
25.8%
540 T
351.9 M
G
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p [bar] T [C] M [kg/s]
Figure 3. An illustration of direct Hot Windbox Repowering
Figure 3 illustrates hot windbox repowering. A gas turbine is installed to provide extra power and
exhausts into the original boiler's windbox. This renders the original air preheater unnecessary, so it
would be removed and the heat it normally recovers from the flue gases before the stack would become
available. This heat can be recovered by new economisers installed in place of the original air preheater.
Those economisers, designated “A” and “B” in Fig. 3, would heat feedwater, which partially bypasses the
original feedwater heaters as described for feedwater repowering (§19.2.1). The new “A” economiser
heats HP water, and the maximum percentage of this water that it can heat is determined by the minimum
pinch allowed between the flue gas and the feedwater. The new “B” economiser heats LP water in an
amount dictated by minimum pinch as well as minimum stack temperature.
The turbine bleed steam is reduced due to lower water flow through the original feedwater heaters. Thus,
steam flow through the LP section of the turbine increases and more power is produced by the steam
turbine in addition to that produced by the new gas turbine(s).
Hot Windbox Repowering is similar in many ways to Feedwater Repowering, discussed in §19.2.1 above.
The main differences are:
1. The Hot Windbox method can allow higher efficiency, since the same air/flue gas goes through both
the gas turbine and original boiler, reducing total plant energy loss at the stack(s).
19-7
 M. A. Elmasri, 1990-2007
Chapter 19: Repowering
2. The Hot Windbox places additional constraints on the selection of the gas turbine(s), since their
exhaust flow rate should be relatively close to the original boiler's airflow (between 75% and 120% of
original airflow, say). Also, the amount of fuel that can be burnt in the boiler is constrained by the fact
that gas turbine exhaust has an oxygen content that is only about two-thirds that of fresh air.
3. The ductwork, windbox and burners of the original boiler will likely need to be replaced, using
materials capable of withstanding the temperature of the gas turbine exhaust, which is significantly higher
than the temperature of the preheated original air. This, along with replacing the air preheater with new
economisers, requires significant downtime of the unit being repowered. The amount of unique
engineering and site work are likely to be costly.
Figs. 4a & 4b show typical results for Hot Windbox repowering with aeroderivative and heavy duty GT
genres. This type of repowering can be done with gas turbine capacities of 30-45% of the original plant,
resulting in a repowered plant capacity that is 110%-150% of original. The repowered plant would have a
net efficiency of about 45-46%, corresponding to a heat rate improvement of about 15-17%.
140
130
%
120
110
Plant net kW
100
ST gross kW
90
Plant net eff.
ST exhaust flow
80
GT Leverage
70
HP FW bypass
Repowering eff.
60
LP FW bypass
50
Boiler fuel
ST inlet flow
Boiler gas
40
30
20
25
29
33
37
41
GT Rating as % of Original Plant Net
Figure 4a. Typical direct Hot Windbox Repowering results with aeroderivatives
19-8
 M. A. Elmasri, 1990-2007
Chapter 19: Repowering
140
130
120
110
Plant net kW
100
ST exhaust flow
ST gross kW
Plant net eff.
%
90
GT Leverage
80
Repowering eff.
70
LP FW bypass
60
Boiler fuel
HP FW bypass
ST inlet flow
Boiler gas
50
40
30
20
26
29
32
38
35
41
44
GT Rating as % of Original Plant Net
Figure 4b. Typical direct Hot Windbox Repowering results with heavy-duty GT’s
With the Hot Windbox scheme, Repowering Efficiency can be very high, well over 100%, at the lower
end of the reasonable capacity spectrum. It should be noted, however, that at the lower end of the
capacity spectrum the cost per incremental kW is quite high, since the GT Leverage is small. The low GT
leverage is due to the derating of the original boiler and steam turbine because of the limited oxygen flow
to the boiler when a gas turbine at the small end of the spectrum is used to supply flue gas to the windbox.
The heavy-duty gas turbine genre is thermodynamically superior to the aeroderivative genre in direct Hot
Windbox repowering, with an advantage of about 0.5 percentage points in net plant efficiency after
repowering. On the other hand, its higher exhaust temperature increases the likelihood that the windbox
and burners have to be rebuilt in a more costly fashion.
19.2.3 HOT WINDBOX REPOWERING WITH FRESH AIR DILUTION
The constraints on designing a direct Hot Windbox repowering may be relieved by blending fresh air with
the gas turbine exhaust before ducting it to the windbox, as shown in Figure 5. This allows greater
freedom to select a smaller gas turbine without starving the boiler of oxygen. It also allows one to cool
the gas turbine exhaust to the point where rebuilding the windbox is unnecessary. Those benefits are
obtained at some expense in efficiency, since diluting the hot exhaust with cool air destroys some exergy.
19-9
 M. A. Elmasri, 1990-2007
Chapter 19: Repowering
366530
8314
43.3
10761
846469
Plant net power
Plant net HR (LHV)
Plant net eff (LHV)
Aux. & losses
Fuel heat input (LHV)
kW
kJ/kWh
%
kW
kWth
Ambient
1.013 p
15 T
60% RH
5.201 p
310963 kW
G
4 LPTs
IPT
HPT
3600 RPM
B
25.31 p
539.6 T
217.8 M
140 p
549.9 T
229.6 M
A
0.04561 p
31.27 T
177.6 M
Stack
124.8 T
333.1 M
13.14 M Fuel
320 M 367.9 T Gas
232.9 T
D
585.4 T
190 M
D
C
A
A
1xKWU V64.3A
66327 kW
D
34.2%
D
B
F
P
B
25.8%
130 M dilution air
G
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p [bar] T [C] M [kg/s]
Figure 5. An illustration of Hot Windbox Repowering with exhaust gas fresh-air dilution
Figure 6 shows typical results. In these calculations, it was assumed that fresh air was to be blended with
the gas turbine exhaust in an amount that reduces its temperature to 245 ºC (473 ºF) before supplying it to
the windbox. Aeroderivatives are thermodynamically superior to heavy-duty gas turbines in this
application, since they require less quenching of their exhaust gases to bring them down to a given
temperature. With aeroderivatives, repowered plant capacity is 100%-120% of original. Aeroderivative
gas turbines for this scheme should be selected at 15-18% of original plant capacity, and should improve
net plant heat rate by about 6-7%. Thermodynamically, heavy-duty gas turbines are less desirable in this
application. If used, they should be selected at 13-16% of original plant capacity and would improve net
plant heat rate by only about 4%.
As a general rule, thermodynamic performance of this scheme is similar, but slightly more efficient than
Feedwater Repowering. Comparison of Figs. 2 and 6 shows that with aeroderivative gas turbines and an
allowable ST exhaust flow increase of 10%, both Feedwater Repowering and Hot Windbox/Dilution
repowering add about the same capacity (20%) to the plant. The Feedwater scheme, however, has a
slightly higher GT Leverage but a slightly lower efficiency. Similar observations apply with the heavyduty gas turbine genre. When comparing the two schemes, however, one should note that Feedwater
Repowering is a less invasive procedure, thus easier to implement. Also, in cases where the ST has
generous exhaust flow and power generation margins, Feedwater Repowering allows taking greater
advantage of them.
19-10
 M. A. Elmasri, 1990-2007
Chapter 19: Repowering
120
110
100
Plant net kW
ST gross kW
90
ST exhaust flow
Plant net eff.
%
80
GT Leverage
Repowering eff.
70
HP FW bypass
LP FW bypass
60
ST inlet flow
Boiler fuel
50
Boiler gas
40
30
20
12
14
16
18
20
GT Rating as % of Original Plant Net
Figure 6. Typical results for Hot Windbox Repowering with fresh air dilution – aeroderivatives
In conclusion, Hot Windbox/Dilution Repowering is not recommended, since the same benefits can be
obtained by the simpler and less costly method of Feedwater Repowering.
19.2.4 HOT WINDBOX REPOWERING WITH EXHAUST PRECOOLING
Another method of avoiding a costly re-build of the windbox and its ductwork is to insert heat recovery
elements in the gas turbine exhaust path before ducting it to the windbox. These heat recovery elements
could include evaporators to generate steam that augments the steam turbine cycle, as well as economisers
to help heat feedwater. The evaporators could be at the main high pressure or at the reheat pressure. In
either case their contribution to the cycle's steam flow is in the range 20-30%. Economisers used to
further cool the gas turbine exhaust before admitting it to the windbox should heat high temperature
water, between the top feedwater heater and the main boiler's economiser. These economisers may also
be configured in parallel with the HP feedwater heaters, bypassing some water around them. Figure 7
illustrates a typical arrangement.
Figure 8 shows that the characteristics of this scheme are generally similar, but somewhat inferior to, the
Direct Hot Windbox Repowering. The reason for this relative inferiority is that the precooling of the gas
turbine exhaust, which is already oxygen-depleted, further reduces the ability of the original boiler to
generate steam. Comparing Fig. 4a with Fig. 8 shows that the capacity gain at 10% LPT overflow is
reduced from about 45% for the direct hot windbox to about 35% with exhaust precooling. GT Leverage
is reduced from about 110% to about 85%. Net repowered plant efficiency is reduced by about 1
percentage point with aeroderivatives and about 1.5 percentage points with heavy duty gas turbines. This
effectively cancels the advantage manifested by the heavy duty genre in direct hot windbox repowering,
so with exhaust precooling, the efficiency of both types is about the same.
19-11
 M. A. Elmasri, 1990-2007
Chapter 19: Repowering
Plant net power
Plant net HR (LHV)
Plant net eff (LHV)
Aux. & losses
Fuel heat input (LHV)
409037
8227
43.76
11980
934778
kW
kJ/kWh
%
kW
kWth
Ambient
1.013 p
15 T
60% RH
5.624 p
319758 kW
HPT
IPT
4 LPTs
G
3600 RPM
B
140 p
484.1 T
259.4 M
A
27.55 p
488.1 T
246 M
0.04998 p
32.89 T
199.5 M
Stack
124.8 T
364.5 M
12.57 M Fuel
351.9 M 305.2 T Gas
255.3 T
To drum
66.5 M
1xKWU V84.2
101259 kW
D
D
A
D
C
A
37.9%
D
F
P
B
B
28.5%
540 T
351.9 M
G
TEC
SEV
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p [bar] T [C] M [kg/s]
Figure 7. An illustration of Hot Windbox Repowering with exhaust gas precooling
140
130
120
Plant net kW
110
ST gross kW
ST exhaust flow
100
Plant net eff.
%
90
GT Leverage
Repowering eff.
80
HP FW bypass
70
LP FW bypass
60
Boiler fuel
ST inlet flow
Boiler gas
50
40
30
20
25
29
33
37
41
GT Rating as % of Original Plant Net
Figure 8. Typical results for Hot Windbox Repowering with exhaust gas precooling – aeroderivatives
19-12
 M. A. Elmasri, 1990-2007
Chapter 19: Repowering
19.2.5 SOME APPLICATIONS OF THE HOT WINDBOX CONCEPT
Feeding gas turbine exhaust, instead of fresh air, into the furnace of a conventional oil-fired or coal-fired
boiler is not new, and the earliest “combined cycles” were built that way. In 1954, such a system was
installed at West Texas Utilities, Rio Pecos, TX, with a GE Frame 3 gas turbine, and in the early 1960’s,
the Arizona Electric Cooperative installed a unit (Apache #1) in which a GE Frame 5 gas turbine
exhausted into the windbox of a B&W conventional boiler.
In The Netherlands, many steam plants have been repowered in this fashion, with original capacities
ranging from 160 to 600 MW before repowering. Additionally, at least one combined cycle was
originally built (in 1982), with the Hot Windbox concept as an original design, just like Arizona Electric
Cooperative’s Apache #1 of the early 1960’s. Most of the repowered plants were originally
commissioned as conventional steam plants in the period 1969-1976, and were re- commissioned after
repowering in the period 1986-1988. All employed heavy duty gas turbines, and most were designed to
improve efficiency without adding appreciable capacity, i.e. with a GT Leverage of near zero. (The paper
ASME 96-GT-250 by Ploumen & Veenema may be consulted for a summary review of the numerous
Dutch hot windbox repowering projects of that period).
In Japan, at least two Hot Windbox repowering projects were implemented in the mid 1990’s by Hitachi.
One plant, originally rated 350 MW, was re-rated 476 MW after installation of a Frame 9E gas turbine in
a Hot Windbox scheme. Another plant, originally rated 700 MW, was re-rated 854 MW after Hot
Windbox Repowering with a Frame 7FA gas turbine.
19.3
AUGMENTING THE EXISTING STEAM BOILER - "PARALLEL REPOWERING"
In this method, a new GT/HRB train is constructed as for a new combined cycle, except without a new
steam turbine or condenser. The HRB is designed to match the existing steam turbine conditions, but
with partial steam production capacity. The rest of the steam for the existing turbine comes from the
original boiler, which is operated at reduced load. Figure 9 illustrates the system.
The main advantages of this scheme are:
1) Flexibility in sizing the additional capacity, typically between 25% and 75% of the original
plant’s.
2) Minimum re-engineering or disruption of the original steam plant during construction of the
add-on.
3) Flexibility of operation, where the original plant can still run when there is an outage in the
new GT system. Likewise, the new gas turbine(s) can run without the steam turbine if the new
plant is fitted with a dump condenser or by-pass stack.
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 M. A. Elmasri, 1990-2007
Chapter 19: Repowering
Plant net power
Plant net HR (LHV)
Plant net eff (LHV)
Aux. & losses
Fuel heat input (LHV)
486159
7729
46.58
14702
1043764
kW
kJ/kWh
%
kW
kWth
Ambient
1.013 p
15 T
60% RH
HP
IP
LP
1xGE 7241FA
Stack
91.32 T
445.4 M
171542 kW
G
329.4 T
44.8 M
547 T
56.57 M
482 T
53.47 M
315.8 T
7.593 M
5.807 p
329319 kW
G
4 LPTs
IPT
HPT
3600 RPM
140 p
549.9 T
247 M
26.98 p
480.7 T
188.5 M
0.05176 p
33.51 T
206.5 M
Stack
112.5 T
231.2 M
11.51 M Fuel
219.7 M Air
232.2 T
D
D
D
C
RE-MASTER NX 5.4 Maher Elmasri Thermoflow
0 09-26-2002 09:45:09 C:\SEMINAR\Sem_5-02\Sec19_Repowering\PAR_GE7FA-3PRH.RNX
D
F
P
HRSG
73.49 M
33.36 T
p [bar] T [C] M [kg/s]
Figure 9. An illustration of Parallel Repowering
Figures 10 and 11 show the performance characteristics of this method, where the new HRB is of the
triple-pressure reheat design. The IP is at the reheat pressure and the LP at the crossover pressure. In
constructing the cases of those figures, it was assumed that the target ST exhaust flow is in the range
100%-110% of its original value; with the upper end of the range used in conjunction with the larger gas
turbine capacities.
Because the new HRB adds more steam at the IP and LP, use of a bigger GT/HRB system requires a
disproportionate turndown in the original boiler and a reduction in HP live steam to the ST inlet, which
operates at reduced kW output based on its exhaust flow limit.
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 M. A. Elmasri, 1990-2007
Chapter 19: Repowering
110
100
90
80
%
70
ST kW
60
ST exhaust flow
50
New HRB HP steam
ST inlet flow
Main Boiler fuel
40
30
20
10
0
20
40
60
80
100
GT Rating as % of Original Plant Net
Figure 10. Typical results for Parallel Repowering with heavy duty GT’s
180
170
160
150
140
130
Plant net kW
%
120
Plant net eff.
110
GT Leverage
Repowering eff.
100
90
80
70
60
50
40
20
40
60
80
100
GT Rating as % of Original Plant Net
Figure 11. Typical results for Parallel Repowering with heavy duty GT’s
Using aeroderivatives in this scheme necessitates some supplementary firing in the HRB to attain the
original steam temperatures. The overall results with both genres of GT are similar; practical gas turbine
capacity being 25%-100% of the original plant rating, resulting in a repowered plant capacity in the range
120%-180% of original. GT Leverage and Repowering Efficiency are both in the 75-85% range for
19-15
 M. A. Elmasri, 1990-2007
Chapter 19: Repowering
either gas turbine genre. The net efficiency of the complete, repowered plant increases from about 43%
with the smallest practical gas turbine capacity to about 50% for the largest practical gas turbine capacity.
This compares with about 39.5% for the original plant.
Parallel Repowering has been implemented a several power plants. Recent examples include Scottish &
Southern’s Peterhead Station in the UK and CFE’s Valle de Mexico Station in Mexico. At Peterhead,
two 660-MW-nominal conventional steam units, originally rated at about 38% efficiency, were repowered
by three Siemens V94.3A, each with a nominal output of 265 MW. The gas turbines feed triple-pressure,
reheat heat recovery boilers that are integrated with the original steam plant. An aerial photograph of the
plant is shown below. At Valle de Mexico, the plant was originally rated 300 MW at 36.8% efficiency as
a conventional steam station. Three Alstom GT11N2 gas turbines were added, allowing it to attain 550
MW at 45% efficiency when operated in Parallel Mode, and to attain 50.6% efficiency at 373 MW
nominal output when operated with GT/HRB’s only, with the original boiler shut down. In the latter
mode, it is more reminiscent of New HRB Repowering, described below.
Aerial view of Peterhead Power Station, repowered with the “parallel repowering” concept in the late 1990’s, with the
two new gas turbines and their heat recovery boilers in the foreground, and the original conventional steam station in
the center (courtesy of Scottish & Southern)
19.4
REPLACING THE EXISTING STEAM BOILER - "NEW HRB REPOWERING"
In this method, a new GT/HRB train is constructed as for a new combined cycle, except without a new
steam turbine or condenser. The HRB is designed with a steam production capacity to suit the existing
19-16
 M. A. Elmasri, 1990-2007
Chapter 19: Repowering
steam turbine. For the reasons explained further below, in many cases the design may call for a lower
throttle flow rate and a proportionately lower throttle pressure, corresponding to operating the steam
turbine with valves wide-open in sliding pressure.
There is an inherent difficulty in repowering an existing conventional steam turbine, designed with
multiple extractions for a feedwater heating cycle, as a combined cycle. In a typical conventional steam
cycle, roughly one-third of the steam is extracted for feedwater heating, leaving two-thirds of the throttle
flow to go to the condenser. Thus, with 100 units of throttle flow, we have 67 units of condensing flow.
A combined cycle works most efficiently without extracting steam for feedwater heating. A single
pressure combined cycle has a condensing flow equal to its throttle flow. A typical 3-pressure reheat CC
has a condensing flow about one-third greater than the throttle flow, i.e. with 100 units of throttle flow, it
has about 133 units of exhaust flow.
Thus, if a conventional steam turbine is repowered as a single-pressure combined cycle, with all
extraction ports plugged, running it with a 100 units of throttle flow would result in 100 units of flow to
the condenser. This is 100/67 = 1.5 times the LPT exhaust design flow. If the conventional turbine is
repowered as a 3-pressure reheat combined cycle with 100 units of throttle flow, it would have 133 units
of flow to the condenser, 133/67 = twice the original LPT exhaust flow. If, on the other hand, the
repowering GT/HRB were sized to match the original LPT exhaust flow, the throttle flow would be twothirds of nominal for a single-pressure, and half of nominal for a triple-pressure reheat configuration.
(Tables 1 and 2 in §14.2.3, which summarise the comparison between turbines for conventional steam
cycles and for combined cycles may be reviewed for additional clarity).
In some plants, the original steam turbine may have some usable flow margin at its low pressure stages,
allowing a condensing flow somewhat greater than the original, and enabling a less drastic HP inlet flow
reduction.
The repowered steam turbine with its reduced throttle flow should still be operated with valves wideopen. Thus, the typical 20-40% reduction in throttle flow will result in operation with a high pressure at
60%-80% of its original value. This is desirable in many cases, where the original HP is too high for
effective heat recovery from typical, unfired GT exhaust.
A high HP is only desirable if the repowered combined cycle were to operate with substantial
supplementary firing, to 750-800 ºC (1380-1470 ºF) say, in which case it should be of single-pressure
design anyway. The single-pressure design will allow a higher throttle flow without overloading the LPT
exhaust and will thus naturally operate at an HP closer to the original value.
Figures 12 and 13 illustrate the characteristics of this type of repowering. The results of three design
alternatives are shown in each diagram and discussed in §19.4.1 through §19.4.3 below.
19.4.1 SINGLE-PRESSURE CC WITH SUPPLEMENTARY FIRING
The rationale for this design is:
(i) Single-pressure to maximise utilisation of the existing steam turbine by minimizing the flow
imbalance between its throttle and its exhaust.
(ii) Minimise the additional gas turbine capacity while still "filling" the original steam turbine,
hence the supplementary firing.
(iii) Using supplementary firing to design with a low HRB stack temperature whilst generating all
the steam at a single, high pressure.
Substantial supplementary firing is required to achieve those objectives (700-820 ºC/1300-1500 ºF). The
numerical examples shown are computed with supplementary firing to 750 ºC/1382 ºF.
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 M. A. Elmasri, 1990-2007
Chapter 19: Repowering
The extreme left curves of Figure 12 show that for this arrangement, the practical range of gas turbine
capacity is 80-120% of original plant output. The upper bound on those figures is based on an assumed
LPT exhaust flow limit at 110% of original. The lower bound is based on obtaining some reasonable,
minimum HP steam flow rate (and pressure), of about 55% of original.
Figure 13 shows that repowered plant capacity would end up in the range 145-200 % of original. GT
Leverage is low at 55-85% due to the necessary derating of the original ST. Repowering efficiency is
quite high, at 70% - 150%. Net efficiency of the repowered plant is about 51.5%, up from about 39.5%
for the original steam plant.
19.4.2 SINGLE-PRESSURE, UNFIRED CC
The rationale for this design is:
(i) As for the design of §19.4.1 above.
(ii) Allow a greater gas turbine capacity than the design of §19.4.1 above.
(iii) The notion that eliminating supplementary firing can increase the efficiency. This turns out
to be untrue with a single-pressure reheat cycle.
The extreme right curves of Fig. 12 show that for this arrangement, the practical range of gas turbine
capacity is roughly 150-200% of original plant output. The upper bound is based on an assumed LPT
exhaust flow limit at 110% of original. The lower bound is based on obtaining at least 55% of the
original HP steam flow.
Fig. 13 shows that repowered plant capacity would end up in the range 210-290 % of original. GT
Leverage is low at 75-90%. Repowering efficiency is 60-70%, lower than the fired cases described above
in §19.4.1.
Net efficiency of the repowered plant is about 51% for the heavy duty genre, up from about 39.5% for the
original steam plant. This is slightly lower than for the fired cases of §19.4.1 above.
19.4.3 TRIPLE-PRESSURE UNFIRED CC
The rationale for this design is to maximise efficiency. Its results appear on the center group of curves in
Figs. 12 and 13.
Fig. 12 shows that for this arrangement, the practical range of gas turbine capacity is roughly 130-150%
of original plant output, with the lower bound based on obtaining at least 50% of the original HP steam
flow. The curves are drawn extending further to the left than is deemed practical.
Comparing the center and rightmost groups of curves on Fig. 12 shows that the gap between inlet and
exhaust flow is wider with triple-pressure, as expected. This exacerbates the imbalance between
insufficient ST inlet flow and excessive LPT exhaust flow, narrowing the practical range of gas turbine
sizes.
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 M. A. Elmasri, 1990-2007
Chapter 19: Repowering
110
100
90
ST kW, 1PRH/750 C
ST ex flow, 1PRH/750 C
ST in flow, 1PRH/750 C
80
ST kW, 3PRH unfrd.
%
ST ex flow, 3PRH unfrd.
ST in flow, 3PRH unfrd.
70
ST kW, 1PRH unfrd.
ST ex flow, 1PRH unfrd.
ST in flow, 1PRH unfrd.
60
50
40
80
100
120
140
160
180
200
GT Rating as % of Original Plant Net
Figure 12. Typical results for New HRB Repowering with heavy duty GT’s
290
270
250
230
Net kW, 1PRH/750C
%
210
Rep. Eff., 1PRH/750 C
190
GT Leverage
170
Rep. Eff., 3PRH unfrd.
150
Net kW, 1PRH unfrd.
Net kW, 3PRH unfrd.
GT Leverage
Rep. Eff., 1PRH unfrd.
130
GT Leverage
110
90
70
50
80
100
120
140
160
180
200
GT Rating as % of Original Plant Net
Figure 13. Typical results for New HRB Repowering with heavy duty GT’s
Fig. 13 shows that repowered plant capacity would end up in the range 200-220 % of original. GT
Leverage is 75-85%. Repowering efficiency is 75-85%.
Net efficiency of the repowered plant is about 54%, up from about 39.5% for the original steam plant.
This is the most efficient type of repowered plant, comparable to a new, green-field, combined cycle.
19-19
 M. A. Elmasri, 1990-2007
Chapter 19: Repowering
19.4.4 RESULTS WITH AERODERIVATIVES
For the single-pressure design with supplementary firing to 750 ºC, discussed in Chapter §19.4.1 above,
aeroderivative results are generally similar, but with a slightly lower net efficiency (about 51% repowered
plant efficiency with aeroderivatives vs. 51.5% for heavy duty gas turbines).
Use of aeroderivatives in the unfired designs of §19.4.2 and §19.4.3 is not feasible, since some
supplementary firing would be needed to attain the proper steam temperatures. With firing to 562 ºC, the
results are very similar to those for unfired heavy-duty gas turbines, except that overall repowered plant
efficiency is lower by about 1% with aeroderivatives.
19.4.5 SOME APPLICATIONS OF REPOWERING WITH A NEW HRB
This type of repowering is the most common. Early US applications include Community Public Service,
Lordsburg, NM (12-MW GE Frame 5K repowering 8-MW ST, 1961); Wheatland Electric Coop, Garden
City, KS (14-MW GE Frame 5L repowering 7-MW ST, 1967). Early European plants include
Drogenbos, Belgium (90-MW Westinghouse 1101 gas turbine repowering 37-MW ST of 1947 vintage,
1976); ESB’s Marina, Cork (repowered 1978 with the a Frame 9B gas turbine) and North Wall, Dublin
(repowered 1983 with a Frame 9E gas turbine). All three of these early European plants used CMI
HRB’s. The early repowering projects involved old steam cycles with low steam conditions and few, if
any, feedwater heaters. Thus, they were well suited to conversion to single-pressure combined cycles
with the relatively modest exhaust flows and temperatures of the gas turbines of the 1960’s and 1970’s.
Recently, many plants were repowered with multi-pressure, reheat HRB’s; examples include the 2001
Senoko, Singapore repowering (3 x Alstom GT26 gas turbines, each with a CMI triple-pressure reheat
HRB replacing the boilers of three 120 MW existing ST’s); and Electrabel’s 2001 repowering at
Vilvoorde, Belgium (1 x Siemens V94.3A + CMI HRB repowering an existing ST).
19.5
OVERALL COMPARISON OF THE VARIOUS REPOWERING SCHEMES
Figs. 14 to 16 provide a succinct summary of the main results discussed above. All three show the plant's
capacity after repowering on the x-axis and the three main "bottom-line" parameters on the y-axis. This
makes it convenient to see what options are available and their comparative characteristics at any given
level of desired additional capacity. In these figures, the following terminology is used to identify the
curves:
-A
-H
FW
WB
WBDIL
WBSEV
PAR
NHRB
3PUF
1PUF
750
960
Aeroderivative
Heavy duty GT's
Feedwater Repowering
Hot Windbox Repowering
Hot Windbox Repowering with exhaust gas fresh air dilution
Hot Windbox Repowering with GT exhaust precooling by a supplementary evaporator
Parallel Repowering
New HRB Repowering
Triple Pressure Unfired
Single Pressure Unfired
Supplementary Fired to 750 ºC (1382 ºF)
Supplementary Fired to 960 ºC (1760 ºF)
If one is repowering the plant to improve efficiency, without increasing capacity, the system of choice is
the Hot Windbox. This approach was widely adopted in The Netherlands during the 1980's, since the
country had a number of modern conventional steam cycles burning natural gas. Repowering Efficiency
is very high for these systems, as seen in Figure 14, but GT Leverage is low, as seen in Figure 15. The
final, repowered plant efficiency is in the low to mid 40's, as seen in Figure 16. The direct hot windbox is
the most efficient and is best with a heavy duty gas turbine. Its drawback is capital cost and the invasive
19-20
 M. A. Elmasri, 1990-2007
Chapter 19: Repowering
nature of the conversion. If one wishes to avoid the cost of rebuilding the windbox, cooling the exhaust
by a supplementary evaporator is much more efficient than cooling it by mixing with fresh air. If the
exhaust is cooled by a supplementary evaporator, there is no tangible efficiency difference between the
gas turbine types.
Repowering Efficiency %
140
130
FW-A
120
WB-A
FW-H
WBDIL-A
110
WBSEV-A
100
WBDIL-H
WB-H
WBSEV-H
90
PAR-A
80
NHRB3PUF-A
70
NHRB750-A
60
NHRB1PUF-H
50
NHRB960-H
PAR-H
NHRB1PUF-A
NHRB3PUF-H
NHRB750-H
40
80
100
120
140
160
180
200
220
240
260
280
300
Repowered Plant Net kW as % of Original Plant
Figure 14. Repowering Efficiency as a function of increased capacity for all repowering schemes
Fig. 14 shows that the notion of Repowering Efficiency loses meaning when the added capacity is small
or non-existent, since it can be negative (if we only add a single kW whilst saving a single BTU of fuel)
or infinite (if we gain any capacity without increasing fuel consumption). It is a very meaningful
quantity, though, as long as the additional capacity is tangible and would need to be obtained by other
means, such as an all new plant, as an alternative to repowering. Repowering efficiency is quite high for
the direct hot windbox. Naturally, this benefit is partly offset by the relatively high capital cost of adding
capacity by this method.
If a modest increase in capacity is sought, by about to 10% - 35% of original capacity, Feedwater
Repowering should be the preferred system. Although the Hot Windbox is also a possibility in this
capacity range, Feedwater Repowering will require less capital cost and downtime. The only argument in
favour of the Hot Windbox would be continued use of a low-grade fuel, such as coal or residual oil, in the
original boiler.
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 M. A. Elmasri, 1990-2007
Chapter 19: Repowering
140
GT Leverage (Net kW gain/GT kW), %
130
FW-A
120
FW-H
110
WB-A
100
WBSEV-A
WBDIL-A
90
WB-H
WBDIL-H
80
WBSEV-H
70
PAR-A
PAR-H
60
NHRB3PUF-A
50
NHRB1PUF-A
40
NHRB750-A
30
NHRB1PUF-H
NHRB3PUF-H
NHRB750-H
20
NHRB960-H
10
0
80
100
120
140
160
180
200
220
240
260
280
300
Repowered Plant Net kW as % of Original Plant
Figure 15. GT leverage as a function of increased capacity for all repowering schemes
Fig. 15 shows that GT Leverage is weak, by definition, when the capacity addition is modest. It is much
weaker with the windbox schemes, where oxygen starvation is the stronger limit, than in the feedwater
scheme, where LPT flow is the governing limit. Thus, situations with a generous LPT flow margin
favour the feedwater scheme.
If a capacity increase of 35% - 70% were sought, one may either use Parallel Repowering or New HRB
Repowering with a single-pressure and supplementary firing (which need to be very heavy towards the
lower end of this capacity range). The former system is more desirable, since it can be accomplished with
less downtime and risk, affords greater operating flexibility, and allows continuing use of low-grade fuel
in the original boiler. The latter is recommended if the original boiler needs to be decommissioned for
other reasons, or if the site cannot accommodate both boilers.
If a capacity increase of 70% - 120% were desirable, then New HRB Repowering is prescribed. The
greatest final efficiency of the repowered plant is achieved with this system, with multiple pressures in the
HRB, without supplementary firing. This efficiency is comparable to that of new, green-field combined
cycles. This system, however, will roughly double the plant's capacity and make it totally dependent on
clean fuel for the gas turbines.
Further capacity increases, up to tripling the plant's original capacity, can be achieved by New HRB
Repowering, with a simple, unfired, single-pressure system. As shown to the extreme right of the
diagrams, this is less efficient than the multi-pressure system.
19-22
 M. A. Elmasri, 1990-2007
Chapter 19: Repowering
Repowered Plant Net Efficiency %
55
54
FW-A
53
52
FW-H
WB-A
WBDIL-A
51
50
49
WBSEV-A
WB-H
WBDIL-H
WBSEV-H
48
47
PAR-A
PAR-H
46
45
44
NHRB3PUF-A
43
42
NHRB1PUF-H
NHRB1PUF-A
NHRB750-A
NHRB3PUF-H
NHRB750-H
NHRB960-H
41
40
80
100
120
140
160
180
200
220
240
260
280
300
Repowered Plant Net kW as % of Original Plant
Figure 16. Final repowered plant efficiency as a function of increased capacity for all repowering schemes
19-23
 M. A. Elmasri, 2006-2007
Chapter 20: Introduction to IGCC
INTRODUCTION TO IGCC
October, 2006, updated November, 2007
© Maher Elmasri 2006-2007
20.1
SYNOPSIS OF COAL UTILISATION TO GENERATE ELECTRICITY
20.1.1 HISTORY OF COAL USE
Coal has been used for heat since, at least, two thousand years ago. There is evidence of coal being
mined in Northeast England in Roman times (Coal Mining & Railways in NE England). By the 13th
century, as England became deforested, wood became less available and expensive, so coal was widely
mined and used for heating. Much of that was so-called “sea-coal”, found along the northeast beaches
where surface coal seams were open to the sea in Northumberland, causing coal to be washed away by
erosion and dumped onto beaches, where it was collected. Coal was also widely mined from open pits
inland, in Durham and Yorkshire since the 13th century. Interestingly, humans have been pollutionconscious for longer than many modern people imagine. Due to its foul smoke, “sea-coal” was banned
for use in London by Edward I in 1272 due to complaints about air pollution. Although violators were
punishable by torture or death, it is said that the ban was widely ignored! i
It is likely that coal surpassed wood as the primary source of energy in England since, roughly, the 16th
century. In America, however, with wide areas of virgin forest, wood was plentiful, and coal surpassed
wood as the most widely used primary fuel only in the late 19th century (ca. 1885, according to the DOE
EIA).
Coal is a “dirty fuel”. Burning coal produces smoke composed of unburnt hydrocarbons and minute
particulates that damage the lungs. Coal contains sulphur which produces noxious gases that damage
many living organisms. When its use as a domestic heating/cooking fuel was still widespread in densely
populated cities, occasional stagnation of the atmosphere resulted in “fogs” that caused many deaths. In
1952 a four-day coal “fog” in London is estimated to have killed 4,000 people. Thus, use of coal as a
domestic fuel has been phased out by laws and regulations in most developed countries. However,
because coal is plentiful and inexpensive, it remains widely used in industry and electric power
generation, but, at least, its use in large facilities provides opportunities to control the pollution it emits.
20.1.2 ELECTRICITY FROM COAL
Coal fired power plants are still the primary source of electricity worldwide. In 2003, coal fired plants
accounted for 40.1% of global electricity production (6.68x 1012 kWh from coal out of 16.66 x 1012 kWh
total). In the USA, coal fired plants produce over half the electricity - 51.5% in 2004 (1.95x 1012 kWh
from coal out of 3.79 x 1012 kWh total).
Coal is likely to remain a primary energy source. The US EIA reports that global reserves of all types of
coal amount to roughly 1012 tons. The current rate of global production is reported as roughly 5.5 x 109
tons. This implies reserves of roughly 180 years’ worth at current production levels. The USA has 27%
of the world’s coal reserves and currently produces 20% of the world’s coal. Thus, US reserves are
equivalent to 240 years’ worth at current production rates.
For the sake of comparison, the US EIA reports 2004 global crude oil reserves at about 1.16 x 1012
barrels, with production at 72.5 million barrels per day. This puts crude oil reserves in 2004 at about 44
years’ worth of 2004 production levels. The estimate given for global natural gas reserves comes out to
i
David Urbinato, EPA Journal, Summer, 1994.
20-1
 M. A. Elmasri, 2006-2007
Chapter 20: Introduction to IGCC
roughly 65 years’ worth at current production levels. Naturally, the accelerating rates of production
would imply much less than 40 years for crude oil and much less than 60 years for natural gas. ii
It should be emphasised, however, that estimates of the world’s fuel reserves tend to be controversial.
Some estimates use the term “proven reserves” (which can only be an estimate anyway, as illustrated by
the recent write-down by the Shell Oil Co.of its own “proven reserves”). Other estimates consider the
rate of new discoveries to project a component of “yet to be discovered” or “probable” fuel reserves.
Thus, pessimists say that the estimates are too optimistic, and the world will run out sooner than thought;
whereas optimists say that estimates are conservative, and the world shall not run out for a long time. In
any case, the consensus is that coal is more abundant than oil or gas, and thus destined to remain the most
commonly used organic fuel in electricity production.
20.1.3 EMISSIONS FROM BURNING COAL
The biggest drawback to producing electricity from coal is emissions. At the present, many old coal
power plants continue operating without any pollution controls whatsoever in many parts of the world.
Particulate control is the highest priority, and is being required for virtually any new coal-fired plant
worldwide, and being retrofitted to many older plants. SOx control is the next highest priority, and this
too is required at an increasing rate around the world, both as a retrofit and as a requirement for new
plants. NOx reduction and CO reduction are lower priorities, but are likely to become increasingly
required in the future. CO2 capture is the lowest priority and can be considered as farthest into the future.
Pollution controls have been retrofitted to many of older coal-fired plants, but at great cost since
engineering a retrofit is always more costly than building the same feature into a new design. Particulate
capture with electrostatic precipitators or fabric filters, flue gas desulphurisation with scrubbers, and
NOx/CO reduction with SCR (Selective Catalytic Reduction) have all been applied to conventional coal
fired plants with success, but at great cost.
The principal pollutants arising from burning coal are summarised below:
20.1.3.1 Particulate Matter (PM):
This includes fine suspended particles of solid dust and liquid droplets of unburnt heavy hydrocarbons.
These fine particles, especially those below 10 µm lodge in the lungs and are injurious to human health.
Coal contains significant amounts of incombustible mineral matter which appears as ash after
combustion. Ash is typically on the order of 10% by weight in many common coals. The range is broad,
with some coals having less than 5% ash, whereas some have more than 30%. In a pulverised coal plant,
a portion of this ash emerges from the boiler as fine particles “flyash” while a portion comes out at the
bottom of the furnace, as slag or “bottom ash”.
Thus, PM emissions depend on the ash content of the coal, and on whether the ash composition and
furnace temperatures are such that a high proportion of the ash fuses and appears as slag, or remains solid
and emerges as dust mixed with the flue gas leaving the boiler.
A 600 MW coal-fired plant, without particulate emissions controls, may emit anywhere from one hundred
to over a thousand tons per day of particulate matter, depending on the type of coal it burns.
Technologies to capture PM are Electrostatic Precipitators (ESP) and Fabric Filters. These are installed
downstream of the boiler, and can capture over 99% of the PM. Virtually all coal fired power plants in
US DOE, Energy Information Administration, Annual Energy Review - DOE/EIA-0384(2004), August, 2005.
The reserve numbers shown are an average of the two sources cited (Oil & Gas Journal and World Oil). These are
stated to be mostly “proven reserves”.
ii
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 M. A. Elmasri, 2006-2007
Chapter 20: Introduction to IGCC
the US have PM controls in place, with ESP accounting for roughly 90% and fabric filters for roughly
10% of installed US coal capacity (of about 300 GW).
20.1.3.2 Sulphur Oxides
Coal contains sulphur, most commonly in the form of pyrite (FeS2). The typical sulphur content in many
of the world’s coals is in the 0.5-1% range. Low-sulphur coals can have as little as 0.2% S by weight, and
high-sulphur coals can contain as much as 4% S by weight. One of the more plentiful coals in the US,
Illinois No. 6, is a high-sulphur coal.
When coal is burnt, the sulphur in the coal is oxidised to sulphur dioxide (SO2). A small percentage of
that SO2, typically between 1% and 10%, depending on excess oxygen and time-temperature history, is
further oxidised to sulphur trioxide (SO3) in the flue gas path before being emitted at the stack.
The SO2 is about two and half times as heavy as air, and sinks to ground level. It has a suffocating odour
and is noxious to humans. It dissolves in rain water, rivers, lakes and oceans, creating sulphurous acid
(H2SO3), which subsequently oxidises to sulphuric acid (H2SO4), harmful to humans, flora, and fauna.
A typical 600-MW-class coal-fired plant burns about 4000 tons of coal per day. If the coal contains only
1% sulphur, and absent any means of reducing sulphur emissions, it spews out about 120 tons/day of
sulphuric acid into the environment.
There are several technologies to reduce SOx emissions. With conventional coal-fired boilers, flue gas
desulphurisation (FGD) is employed downstream of the boiler, before the stack. There are several
methods of FGD, the most widely used being the wet scrubber employing a shower of limestone/water
slurry sprayed through the flue gas in a tank. The limestone (CaCO3) reacts with SOx to produce gypsum
(CaSO4), which may also be sold as a byproduct. There is also a dry systems that employ lime (CaO), as
well as a system that simply scrubs the flue gas with seawater, dissolving much of the SOx, then dumping
the dilute acid thus formed back into the ocean.
FGD is mandatory for new power plants in most parts of the world. Adding FGD to existing plants is
costly, but being phased in as a retrofit in many parts of the developed world. Roughly 100 GW of the
US’s 300 GW of coal-fired power generation capacity (i.e. about one-third) has FGD. By 2015, it is
expected that 180 GW (i.e. about 60%) of the US’s coal-fired power generation capacity will have FGD.
Instead of absorbing SOx post-combustion from the flue gas, one proven technology is to absorb it during
combustion in a fluidised bed boiler. By mixing pulverised limestone with the coal particles and
suspending the mixture in the combustion air stream, the sulphur is captured and converts a portion of the
limestone to gypsum. This technology can offer a more cost-effective approach to capturing sulphur than
using wet limestone scrubbers, particularly for smaller units (below 200 MW). Fluidised bed boilers are
not common in the US, but are widely used in some parts of the world, most notably in China.
20.1.3.3 Nitrogen Oxides
The term NOx is commonly used to represent the sum of the two principal nitrogen oxides found in the
atmosphere, nitric oxide (NO), and nitrogen dioxide (NO2), which react with each other and with the
oxygen in the atmosphere creating ozone. There are several other forms of nitrogen oxides, like nitrous
oxide (N2O), the “laughing gas” once used as an anaesthetic, and N2O3, N2O4, and N2O5. NOx is harmful
to human health, and contributes to acid rain since it eventually creates nitric acid in the environment.
NOx results from high-temperature combustion, and is difficult to calculate with precision because its
formation depends on the excess air, the local temperature-time history within the flames, and whether the
coal contains significant amounts of fuel-bound nitrogen in addition to the nitrogen present in the
atmospheric air used for combustion. Although the range of NOx emissions from coal plants varies
widely, the order of magnitude for a 600 MW plant is about 30 tons/day.
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Chapter 20: Introduction to IGCC
Adding SCR to existing coal fired plants is costly, but being phased in. Roughly one-quarter of the US’s
conventional coal power generation capacity now has SCR. By 2015, it is expected that just over half the
US’s conventional coal-fired power generation capacity will have SCR.
20.1.3.4 Carbon Monoxide
This noxious pollutant results from non-equilibrium in the combustion process, so, like NOx, it is difficult
to calculate with precision. The rough order of magnitude expected from a 600 MW pulverised coal plant
is about one ton per day.
20.1.3.5 Heavy Metals
Coal ash contains traces of harmful heavy metals, chiefly mercury. Typical US coals contain mercury in
the range of 0.1 ppm. Although the concentration of mercury and its oxides in flue gases is in the parts
per billion, there is new emphasis on its control. Fortunately, PM controls and wet FGD scrubbers are
also effective in reducing mercury emissions.
20.1.3.6 Carbon Dioxide
CO2 is not considered a pollutant, but a greenhouse gas. There is much controversy, political and
scientific, about whether it really contributes to global warming, or even whether global warming is itself
real. However, sufficient credible scientific opinion has weighed in on the probability that the increasing
levels of CO2 in the atmosphere are causing irreversible changes to the environment that the world’s
developed nations, except the USA, have agreed to limit its discharge into the atmosphere. Many leaders
in industry and government in the US believe that regulations to limit CO2 emissions are likely to be
enacted in the US within the next decade.
Laws to control CO2 emissions and setup trading of CO2 allowances have been implemented, or are in
process, in many industrialised countries. The trading price of these allowances is currently in the range
$12-$25/ton. This price is expected to rise to levels that make CO2 capture, for use in enhanced oil
recovery, other industries, or even merely for sequestration in appropriate geological formations; an
increasingly profitable business. It should be mentioned that capturing flue gas CO2 and using it for
enhanced oil recovery is already viable economically in certain situations, and was even viable and used
profitably since the late 1970’s.
As may be expected, the rates of discharge of CO2 from a coal-fired plant are large, and may be
accurately calculated on the basis that essentially 100 % of the carbon in the coal is oxidised to CO2, so its
production rate is simply a function of the coal’s carbon content and heating value. A typical 600 MW
coal-fired plant will produce roughly 12,000 tons per day of CO2.
Regardless of if and when CO2 emissions will be limited, or CO2 sequestered, all of us can contribute to
limiting CO2 emissions by simply avoiding the wanton waste of energy. Each mile of unnecessary
driving produces about 1 lb of CO2. Each kWh of electricity we waste in the US corresponds to roughly 2
lbs of incremental CO2 emissions. A common example of electricity wastage is leaving a personal
computer on all the time. An idling PC, even with its monitor on standby, consumes about 75 W of
power. If it is actually used 1760 hrs per year, but left on unnecessarily for the balance of the time (7000
hrs each year), it is wasting 525 kWh and causing wanton emissions of about ½ a ton of CO2 per year.
When you consider that millions of PC’s are left on all the time all over the USA, you can see that
millions of tons of CO2 are wantonly emitted just by neglect of the simple measure of turning off a PC at
the end of each working day.
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Chapter 20: Introduction to IGCC
20.2
COAL GASIFICATION
20.2.1 HISTORICAL
Gasification is the process of extracting a gaseous fuel from a solid one, such as coal or wood. The
simplest and oldest method is to heat the solid fuel in a retort in the absence of air, driving out the volatile
hydrocarbons and leaving a solid char, mostly carbon. The first commercial-scale gasification industry
was started by Philippe Lebon in France at the end of the eighteenth century.
Most coals (except anthracite) contain a substantial percentage of volatile matter, typically between 20%
and 40%; and wood may contain up to 75% volatile matter. These volatiles evaporate as the coal or wood
is “cooked”, and in early gasification schemes these vapors were allowed to flow through a water bath to
clean them up. The heavier volatiles condensed in the water bath, creating a “liquor” of heavier
hydrocarbons and water, whereas the lighter components passed through the bath without condensing,
and this “clean” gas was then piped to the end-user. This gas produced a luminous flame when burnt, and
was used for street lights in many cities in the nineteenth and early twentieth century. The residue left in
the retort, the coke, with a high carbon content, was also used either as a solid fuel or as a reducing agent
in the extraction of iron from ore. The liquors resulting from washing the gas, which contain many
harmful tars and heavy hydrocarbons, were frequently dumped into the environment.
In the late nineteenth century, the manufacture of “water-gas”, which had a higher calorific value was
developed and widely used. This process involved passing steam over hot coal or coke in the absence of
air, producing a mixture consisting largely of hydrogen and carbon monoxide (C + H2O → CO + H2),
with small amounts of methane, carbon dioxide, nitrogen, and sulphur compounds. This syngas burns
without much luminosity, and so was used for heating and cooking rather than gas lamp lighting. Due to
the high content of carbon monoxide, this syngas was highly poisonous, but, nonetheless, was widely
piped as a domestic fuel in cities until it was replaced by natural gas in recent times.
Modern gasification is similar, except that rather than heating the coal in the absence of air by external
means, as in the nineteenth century methods, the coal is partially combusted with insufficient air to
generate the high temperature needed for gasification. This allows a continuous process, rather than the
early retort methods which required a batch process in a closed retort.
To keep the size of the equipment compact, the modern gasifier is a vessel operated at high pressure,
wherein coal and oxidant are introduced continuously, with partial combustion taking place and syngas
being produced at high temperature at the vessel’s outlet.
20.2.2 HOW MODERN GASIFICATION FACILITATES EMISSIONS CONTROL
In conventional coal fired power plants, emissions are dealt with downstream of the boiler. The coal is
burnt as-is, heat is extracted from the combustion products, then any pollution control measures are
applied to the flue gas stream after it leaves the boiler. This approach has the advantage that pollution
controls can be applied optionally, and as an afterthought, downstream of an already well-established
technology dating back to the time when emissions were of no concern. It has the disadvantage that it can
be very expensive since the volume of flue gases is large, and since the variety of pollutants in the flue
gas are all present and have to be dealt with at once.
If the coal is first gasified, and the resulting syngas used to run the power plant, the emissions control
problem is simplified because the pollutants are concentrated in a compact syngas stream, rather than
dispersed in voluminous flue gases. For the same net plant power output, the syngas stream in an IGCC
has about 15% of the mass flow rate of flue gases leaving an equivalent conventional steam cycle boiler.
Because the syngas is pressurised, its volumetric flow rate is roughly 1% that of the flue gas leaving an
equivalent conventional steam cycle boiler. This enables more compact and cost-effective emissions
control equipment.
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Chapter 20: Introduction to IGCC
20.2.3 MODERN GASIFIER TYPES: OXYGEN-BLOWN & AIR-BLOWN
20.2.3.1 Oxygen-Blown Gasifiers
An oxygen-blown gasifier uses oxygen, instead of air, for the partial combustion and gasification of the
fuel. Thus, it produces syngas undiluted with nitrogen, with a higher calorific value. This allows the gas
processing equipment to be more compact, and, if the resulting syngas is to be burnt in a gas turbine, the
gas turbine will need fewer modifications to its combustor and turbine flow path. On the other hand, this
system requires an air-separation plant, which consumes a considerable amount of power, on the order of
7-10% of the gross output of such an IGCC. The air separation plant also adds cost and complexity, and
detracts from reliability.
Syngas properties depend on coal composition and many variables in the process, but as a guideline, an
oxygen-blown gasifier will produce syngas of the following approximate properties (after cleanup):
Table 1. Typical Dry, Clean Syngas Composition from Oxygen-Blown Coal Gasifiers
Property
Typical Value
Typical Range
Carbon monoxide (CO) Volume %
52%
40%-60%
Hydrogen (H2) Volume %
32%
25%-40%
Carbon dioxide (CO2) Volume %
12%
4%-16%
N2, H2O, CH4, etc, Volume %
4%
3%-6%
Molecular weight
21.5
20-23
Volumetric LHV kJ/scm
9,300
8,000-10,000
250
215-270
LHV kJ/kg
10,500
8,200-11,600
LHV BTU/lb
4,500
3,500-5,000
Volumetric LHV BTU/scf
Amongst the modern oxygen-blown gasifiers developed for IGCC are the GE (former Texaco), Shell, and
E-Gas (Conoco Philips, former Destec), and Siemens (former Sustec) gasifiers.
20.2.3.2 Air-Blown Gasifiers
Using air as the oxidant simplifies the system by dispensing with an air separation plant. However, the
syngas fuel so produced is diluted with atmospheric nitrogen, so it has an even lower calorific value than
syngas produced from an oxygen-blown gasifier.
Although syngas properties depend on coal composition and many variables in the process, as a guideline,
an air-blown gasifier will produce syngas of the following approximate properties (after cleanup):
Table 2. Typical Dry, Clean Syngas Composition from Air-Blown Coal Gasifiers
Property
Typical Value
20-6
Typical Range
 M. A. Elmasri, 2006-2007
Chapter 20: Introduction to IGCC
Nitrogen (N2) Vol %
58%
55%-62%
Carbon monoxide (CO) Vol %
23%
20%-25%
Hydrogen (H2) Vol %
13%
11%-15%
Carbon dioxide (CO2) Vol %
5%
3%-7%
Ar, H2O, CH4, etc Vol %
1%
0.5%-2%
Molecular weight
25.5
25-26
Volumetric LHV kJ/scm
3,900
3,500-5,500
105
95-150
LHV kJ/kg
3,750
3,400-5,300
LHV BTU/lb
1,600
1,500-2,300
Volumetric LHV BTU/scf
Amongst the modern air-blown gasifiers developed for IGCC are the Mitsubishi and the TRIG
(KBR/Southern Co.).
It should be added that some “air-blown” systems actually incorporate a small air separation plant, and
enhance the gasifier air by adding some oxygen, whilst using some of the nitrogen as a transport gas for
the dry coal feedstock, and, perhaps, injecting some of the surplus nitrogen into the gas turbine. With
oxygen enhancement, these gasifiers will produce syngas towards the upper end of the range of heating
values given in Table 2. Naturally, the heating value may exceed the upper range given in Table 2 if the
additional oxygen is considerable, since, in the limit, the heating value will approach that of oxygenblown systems.
20.2.4 MODERN GASIFIER TYPES: SLURRY FEED & DRY FEED
Since the gasification vessel is at high pressure (typically 30-60 bar, i.e. 450-900 psia), the coal feedstock
has to be pumped in.
One method is to mix the pulverised coal with water (roughly water/coal ratio of about 0.5 by weight) to
create a slurry, which can be pumped and controlled much like a fluid. Slurry pumps, though, are
specialised units, and a typical design will be a positive displacement type with a reciprocating
piston/diaphragm system.
Since the gasification chemistry is enhanced by the presence of H2O to produce hydrogen, the slurry feed
system doubles as a method of introducing the H2O. However, because the H2O is introduced as liquid, it
reduces IGCC efficiency because the latent heat to evaporate it comes from fuel, rather than from lowgrade heat recovery if the H2O were introduced as steam that had been generated in a HRSG, syngas
cooler, or other low-grade heat recovery device.
In dry feed systems the pulverised coal is introduced as dust, transported by an inert gas like nitrogen.
This is commonly done through a batch process with lock hoppers. An atmospheric-pressure hopper is
filled with coal, sealed, then pressurised with nitrogen. It may then discharge its contents into the
gasifier, while another lock hopper is being filled and pressurised.
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Chapter 20: Introduction to IGCC
With dry feed systems, the H2O is typically introduced as steam, which is evaporated using low-grade
heat, helping to boost IGCC efficiency.
Amongst the modern gasifiers developed for IGCC, the GE and Conoco-Philips gasifiers uses slurry feed,
whereas Shell, Siemens, Mitsubishi, and KBR/Southern Co. use the dry feed system.
20.2.5 MODERN GASIFIER TYPES: SINGLE-STAGE VS TWO-STAGE
In a single stage gasifier, the coal is injected into the oxidant stream, which has insufficient oxygen for
complete combustion, and partial combustion to generate high temperatures and gasification thus occur
simultaneously.
In a two-stage concept, a small amount of coal (and/or recirculated char) is injected into an oxidant stream
that contains enough oxygen to burn completely. This primary zone generates very hot flue gases. The
coal to be gasified is then injected into these hot flue gases in the secondary zone (also called reducing
zone), where it gasifies and cools the gases. Any excess oxygen leaving the primary zone is consumed in
partial combustion within the secondary zone. Two-stage gasifiers tend to be more efficient, with a
greater proportion of the coal’s energy emerging in the form of chemical calorific value in the syngas, and
a smaller proportion as sensible heat in the hot syngas.
Amongst the modern gasifiers developed for IGCC, the GE, Shell and Siemens gasifiers are single-stage,
whereas Conoco-Philips, Mitsubishi, and KBR/Southern Co. use two-stage systems.
20.3
MODIFYING A GAS TURBINE TO BURN SYNGAS
Pure methane has an LHV on a volumetric basis of 32,824 kJ/scm (881 BTU/scf) and on a mass basis of
50,047 kJ/kg (21,516 BTU/lb). Typical natural gas has a calorific value that is slightly less.
Syngas has a much lower calorific value, both on a mass basis and on a volume basis, than natural gas.
Since standard gas turbine models were designed to burn natural gas, some modifications to the
combustion system and turbine flow path may be necessary.
20.3.1 COMBUSTION SYSTEM MODIFICATIONS
The combustion system must handle a higher volumetric flow rate of syngas to achieve the same heat
input as with natural gas.
After processing by cleaning and drying, the syngas produced from a typical oxygen-blown gasifier will
have a calorific value per unit volume that is roughly 1/3rd that of natural gas. Syngas from an air-blown
gasifier (without oxygen enhancement) will have a calorific value per unit volume that is roughly 1/8th
that of natural gas. This implies syngas volumetric flow rate to the gas turbine about triple that of natural
gas with oxygen-blown gasifiers, and about eight times that of natural gas with air-blown gasifiers.
Naturally, these large multiples of fuel volumetric flow rate require substantial re-engineering of the gas
turbine’s burners, fuel delivery and fuel control systems.
20.3.2 TURBINE FLOW PATH & FIRING TEMPERATURE CONTROL MODIFICATIONS
As mentioned elsewhere in this seminar, the GT cycle pressure ratio is set by the flow rate and properties
of hot gases entering the turbine’s first stage nozzles.
Since syngas is denser than natural gas, syngas mass flow rate multiples are even higher than the
volumetric flow rate multiples given above. Mass flow rate of oxygen-blown syngas is about five times
that of natural gas, and mass flow rate of air-blown syngas is about thirteen times that of natural gas. This
large additional fuel mass enters the turbine after combustion, increasing its mass flow rate.
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 M. A. Elmasri, 2006-2007
Chapter 20: Introduction to IGCC
Thus, in a typical oxygen-blown IGCC, if the gas turbine compressor delivers its standard airflow to the
combustor, and if the GT is operated at the same firing temperature, maintaining the standard pressure
ratio would require an increase in the throat area of the turbine’s first-stage of about 8% to accommodate
the extra exhaust gases resulting from the syngas. Under such conditions, the exhaust mass flow rate
would be about 9% greater than with natural gas, and the exhaust temperature would be about 30° F (17°
C) higher than with natural gas. This increase in exhaust temperature is primarily due to the fact that the
turbine gas flow rate is higher, but the turbine cooling air flows remain unchanged, so the quenching
effect of the turbine cooling air on the hot gas path is weakened. The diminution of turbine cooling air
flow in proportion to the flow of the expanding hot gases, however, will tend to overheat the turbine
blades. Thus, either an increase in turbine cooling air flow or a reduction in firing temperature is in order.
The latter solution is more conservative, and is further discussed below.
From the above, we observe that to maintain the design compressor flow, firing temperature, and pressure
ratio with oxygen-blown syngas, we need to increase turbine throat areas by 8%, as well as increase the
exhaust temperature control set-point by about 30° F (17° C), but all this will overheat the turbine blades
relative to natural gas. Alternatively, we can maintain the exhaust temperature at its normal natural gas
set point, which, in effect, lowers the corresponding firing temperature by about 45 °F (25 °C) relative to
natural gas. The lowered firing temperature will tend to lower the pressure ratio. Thus, if the exhaust
temperature set point is left unchanged, effectively derating the firing temperature by about 45 °F (25 °C),
maintaining the design pressure ratio will only require a 6% increase in turbine throat areas, rather than
the 8% described above if the firing temperature were not derated.
20.3.3 INTEGRATING THE GAS TURBINE CYCLE WITH THE GASIFIER
Some oxygen-blown IGCC schemes involve some integration between the ASU and the gas turbine cycle.
This can be in the form of bleeding off some compressed air from the GT compressor and sending it to
the ASU. The rationale is that this saves power in the ASU compressors, and ameliorates the problem of
excess mass flow through the turbine by diverting some of the GT compressor’s air.
This approach has many drawbacks:
(a) The ASU operates at a pressure much lower than that at the GT compressor discharge. A typical
ASU compressor train delivers air at about 4-6 bars (60-90 psia), whereas a modern large GT
compressor delivers air at about 16-20 bars (240-300 psia). This means either taking the air off
the compressor at an intermediate bleed, thereby upsetting the compressor’s normal airflow
distribution, or taking the air at the compressor’s discharge and wasting it excess energy, both
pressure and temperature. Attempting to take the air from the compressor discharge and
recovering energy from it before sending it to the ASU would create excessive complexity.
(b) Such integration complicates design, plant layout, and construction, since the ASU is normally
provided by companies that specialise in production of industrial gases, whereas the power island
is designed by companies that specialise in power plants.
(c) Such integration increases complexity of operation, particularly during start-up and shut-down.
Another form of integration in oxygen-blown IGCC schemes is to inject all or part of the nitrogen
produced by the ASU into the gas turbine, at the compressor discharge or at the combustor. This has the
opposite effect of bleeding air from the compressor, i.e. it increases the mass flow rate through the
turbine, thereby requiring an even greater increase in turbine throat areas. There are two reasons for
injecting N2 into the GT. First, to reduce NOx by diluting the fuel in the combustion zone to lower peak
flame temperatures. This is a sensible approach to the problem of NOx abatement. The second reason
sometimes given is to use the compressed nitrogen available from the ASU for power augmentation,
rather than waste it. This rationale is dubious, because the N2 leaving an ASU is normally at modest
pressure, say about 2-3 bars (30-45 psia), and thus requires very considerable power to compress it to a
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 M. A. Elmasri, 2006-2007
Chapter 20: Introduction to IGCC
pressure that must be above that of the GT combustor by 20%-65% to be able to control its flow rate and
introduce it into the GT combustor system.
For the reasons given above, it is the opinion of the present author that it is best to avoid integration of the
ASU with the gas turbine. It is best to have the ASU function separately, possibly with multiple trains.
The ASU can be engineered for maximum energy efficiency by having its own systems for compressing
air and recovering power from its own unwanted N2 by-product.
20.4
A FEW EXAMPLES OF OXYGEN-BLOWN IGCC PROCESSES
20.4.1 “GE PROCESS” EXAMPLE
The diagram below shows the overall schematic of an oxygen-blown IGCC of 630 MW net output (714
MW gross less 84 MW of auxiliaries). It burns a high-sulphur Eastern coal, and is based on two GE
Frame 7FB gas turbines integrated with a gasification scheme that includes “full heat recovery” from the
syngas prior to its cleanup. Overall coal-pile to bus-bar efficiency is 44.04% based on LHV (41.24% on
HHV). A plant of this sort is expected to cost roughly $2000-$2400/kW in 2007 dollars. About one-third
of the total cost is for the power island proper, and two-thirds of the cost for the coal gasification side.
Gross Power = 714292 kW, Net = 630333 kW
LHV Gross Heat Rate = 7214, Net = 8175 kJ/kWh
LHV Gross Electric Eff. = 49.9 %, Net = 44.04 %
Scrubber
IGCC System Block Flow Diagram (Plant totals of 2 gasifier trains)
- Type 1 Gasifier with Radiant and Convective Coolers
33.32 p
178.5 T
134 M
COS
Hydrolysis
33.32 p
179 T
134 M
33.32 p
357.2 T
119.9 M
35.08 p
732.2 T
119.9 M
Steam
130.9 p
125.5 M
RSC1
(Radiant)
36.92 p
1371.1 T
119.9 M
Slag
5.858 M
Gasifier
Type 1
Cooler 1
Pittsburgh No. 8
25 T
51.71 M
4468 tonnes / day
Ambient air
1.013 p
15 T
880.7 M
Cooler 3
30.52 p
37.78 T
99.73 M
30.52 p
37.78 T
98.16 M
AGR
ASU
H2S
1.571 M
Discharge
161 M
Oxygen
36.92 p
113.4 T
51 M
Water
23.07 M
30.75 p
79.44 T
100.8 M
Ambient air
1.013 p
15 T
212 M
631.2 T
978.9 M
Syngas
29.99 p
174 T
98.16 M
Slurry
74.78 M
Fuel
Preparation
Cooler 2
29.99 p
174 T
98.16 M
Steam
130.9 p
65.11 M
RSC2
(Convective)
32.8 p
169.3 T
125.8 M
18.65 p
416.9 T
17.72 p
1344.7 T
HRSG
124 p
566 T
3516 h
214.9 M
1.048 p
632.2 T
489.4 M
321367 kW
392925 kW
ST
GE 7251FB
(Totals of 2 units)
p [bar], T [C], h [kJ/kg], M [kg/s], SteamProperties: Thermoflow - STQUIK
GT PRO 16.0 Mims
0 09-26-2006 11:11:04 file=C:\Tflow16\MYFILES\SEM01.GTP
Figure 1. Example of a “GE Process” IGCC with full raw syngas heat recovery
20-10
Stack
1.013 p
151.8 T
978.9 M
0.0689 p
38.74 T
2445 h
214.7 M
 M. A. Elmasri, 2006-2007
Chapter 20: Introduction to IGCC
20.4.2 HEAT RECOVERY FROM THE RAW SYNGAS
The plant model summarised above cools the raw syngas prior to scrubbing by recovering heat from it in
two stages, first a radiant raw syngas cooler (RSC1) cools it from the gasifier exit temperature of 1371 °C
(2500 °F) to 732 °C (1350 °F), then a convective raw syngas cooler (RSC2) further cools it to 357 °C
(675 °F).
The mass flow rate of raw syngas is about 12% of the GT exhaust mass flow rate. The syngas heat
recovery available temperature drop is about 1000 °C (1800 °F), double the temperature drop of flue
gases across the HRSG, which is on the order of 500 °C (900 °F). Hot raw syngas has a higher specific
heat than cooler HRSG gases, so the raw syngas coolers have a heat recovery potential on the order of 2025% of the heat recovery potential of the HRSG.
Because raw syngas heat is available at high temperature, it is best employed to make HP steam. To
make sure the raw syngas heat exchanger surfaces don’t overheat, they are employed as evaporators,
without economisers or superheaters. This is because boiling heat transfer coefficient is very high in an
evaporator, keeping metal temperatures close to saturation. Thus, in an IGCC, the water going to the
syngas coolers is preheated in the HRSG’s economisers, and the resulting steam is superheated in the
HRSG’s superheaters. This creates an unusual HRSG temperature profile, illustrated in Fig. 2, where the
HRSG has little evaporation, but high duty in its economisers and superheaters. In this example, about
11% of the HP economiser water flow rate is evaporated in the HRSG and 89% in the raw syngas coolers
(59% in the radiant, 30% in the convective). The HP economisers absorb so much heat, that there is none
left for an IP evaporator, and very little for an LP evaporator. Thus, this type of IGCC needs a singlepressure or dual-pressure reheat HRSG, but not triple-pressure reheat.
GT PRO 16.0 Mims
700
LPB
9282
HPE2
33208
HPE1
41559
Net Power 630333 kW
LHVHeat Rate 8175 kJ/kWh
HRSG Temperature Profile
HPB1
14106
HPE3
20078
LPS
HPS0
50.57
62789
HPS1
17735
RH1
34941
RH3
14250
HPS3
11295
0
1
2
Q kW
600
3
4
TEMPERATURE [C]
500
400
5
7
8
10
300
12
200
14
UA kW/C
100
531.9
1996.6
1859.6
0
0
50
0.68
935.6
1841.9
426
452.7
100
150
200
240.5
248.2
155.6
250
300
HEAT TRANSFER FROMGAS[.001 XkW] (per HRSG)
Figure 2. HRSG profile for the cycle of Fig. 1, with full raw syngas heat recovery
The radiant raw syngas cooler is similar to a furnace in a conventional PC plant, with membrane
waterwalls to receive the radiant heat flux, but differs from a conventional furnace in that it contains high
pressure hot gases, at 30-40 bars, instead of gases at near-atmospheric pressure. This makes the radiant
20-11
 M. A. Elmasri, 2006-2007
Chapter 20: Introduction to IGCC
syngas cooler much smaller than an equivalent conventional furnace, since the dense gases require a
smaller flow cross-section and have higher emissivity. The radiant cooler is generally in the form of a
long thin cylinder, as illustrated in Figure 3, with thick walls to withstand the pressure, and with
insulation between the outer pressure casing and the inner membrane walls.
7
6
5
4
3
2
1
FOR QUALIT AT IVE INDICAT ION ONLY
8
A
A
D
A
B
B
C
C
C
B
E
D
D
E
E
Thermoflow, Inc.
Company: Thermoflow, Inc.
User: Mims
GASIFIER
F
A
B
C
D
E
F
G
H
I
J
10.8 m
46.3 m
60.6 m
3.9 m
5.2 m
-
-
-
-
-
1
2
3
5
4
ELEVATION VIEW
Date: 11/01/06
F
Drawing No:
C:\Tflow16\MYFILES\SEM01.GTP
6
7
8
PEACE/GT PRO 16.0.1 Mims
11-01-2006 10:53:33
Figure 3. Rough dimensions of an oxygen-blown gasifier (top cylinder) and radiant raw syngas cooler (bottom
cylinder), sized for a 200-MW-class gas turbine
The convective raw syngas cooler used in some pilot plants is similar to a water-tube boiler, with hot,
pressurised flue gases flowing within tubes immersed in a pool of boiling water. The tubes are typically
about 2.5” (57 mm) in diameter. This arrangement results in a compact design, but because the
contaminated gases flow in tubes which have wall temperatures cool enough for ash and corrosive
components to condense & solidify, these designs have been subject to fouling, corrosion, and deposition
within the tubes. As a result of this experience, some of the suppliers of IGCC are shying away from
using a convective cooler in the next generation of IGCC, and simply quenching the raw syngas with
water after the radiant cooler.
If the syngas is quenched after the radiant cooler, there is less HP steam generated, so the HRSG
economisers and superheaters have less duty to heat water and superheat steam that is evaporated
externally to the HRSG. This results in a HRSG temperature profile “closer to normal”, as shown in
Figure 4. Here, about 32% of the HP economiser water flow rate is evaporated in the HRSG and 68% in
the radiant raw syngas cooler. The HP economisers, however, still absorb so much heat, that there is
insufficient low-grade heat for an IP evaporator, although the LP evaporator becomes prominent. Thus,
this type of IGCC needs a dual-pressure reheat HRSG, but not triple-pressure reheat.
Naturally, eliminating the convective syngas cooler lowers the IGCC power and efficiency, because less
HP steam is available to the steam turbine. For the cycle example illustrated by Fig. 1, plant net output is
reduced by about 6%, from 630 to 594 MW; and net plant efficiency drops from 44.04% to 41.65%.
20-12
 M. A. Elmasri, 2006-2007
Chapter 20: Introduction to IGCC
GT PRO 16.0.1 Mims
700
HPE1
35705
LPB
20331
Net Power 594332 kW
LHVHeat Rate 8644 kJ/kWh
HRSG Temperature Profile
HPE2
HPB1
28531
34356
HPE3
17250
LPS
2320.7
HPS0
53833
2
Q kW
600
HPS1
15206
RH1
29803
HPS3
9684
RH3
12196
0
1
3
4
500
TEMPERATURE [C]
5
400
7
8
10
300
12
14
200
UA kW/C
100
825.5
994.4
1234.5
0
0
50
28.95
710.9
868.4
776
100
150
306.7
200
191.5
190.1
131
250
300
HEAT TRANSFER FROMGAS[.001 XkW] (per HRSG)
Figure 4. HRSG profile for the example of Fig. 1, but with syngas quench replacing the convective raw syngas
cooler, reducing the amount of HP steam evaporated in the syngas heat recovery
In the quest for reliability and simplicity, some IGCC designs eliminate all heat recovery from the raw
syngas and quench it right after the gasifier. This further reduces output and efficiency. For the cycle
example of Fig. 1, if all syngas heat recovery were eliminated, plant net output would fall by about 17%,
from 630 to 526 MW; and net plant efficiency would drop from 44.04% to 36.94%, calling into question
the entire notion of IGCC.
20.4.3 “SHELL PROCESS” EXAMPLE
This is shown in Figure 5. It differs thermodynamically from the “GE Process” in two ways: (a) Dry
feed, rather than slurry feed, and (b) greater heat recovery from the raw syngas. Dry coal feed can
improve efficiency, as described earlier, and, naturally, recovering more heat from the raw syngas before
scrubbing also contributes to efficiency, but at the cost of having two convective syngas coolers, one
making HP steam, and another making IP steam. The net result is an efficiency improvement of about
one percentage point with these features, i.e. about 45.2% net LHV efficiency in the example of Figure 5,
vs. 44.04% in the example of Figure 1. Both these examples have otherwise similar assumptions.
This process avoids the radiant syngas cooler altogether by recirculating cooled syngas to quench the hot
gases leaving the reaction zone of the gasifier. This does not incur any additional thermodynamic
penalty, other than recirculating fan power, because the radiant cooler cannot take full advantage of the
excess temperature (above HP saturation temperature) of the very hot gasification products anyway.
Thus, as long as the quenched syngas in this process is still well above HP saturation temperature, no
additional exergy loss is introduced.
20-13
 M. A. Elmasri, 2006-2007
Chapter 20: Introduction to IGCC
Gross Power = 674228 kW, Net = 599030 kW
LHV Gross Heat Rate = 7074, Net = 7962 kJ/kWh
LHV Gross Electric Eff. = 50.89 %, Net = 45.21 %
Scrubber
Recir.
105.3 M
33.32 p
242.9 T
201.7 M
35.08 p
898.9 T
201.7 M
Steaminj
130.9 p
4.787 M
Pittsburgh No. 8
25 T
47.87 M
4136 tonnes / day
- Type 2 Gasifier with Convective Cooler
33.32 p
145.7 T
102.3 M
RSC2
(Convective)
Slag
5.09 M
IGCC System Block Flow Diagram (Plant totals of 2 gasifier trains)
Gasifier
Type 2
Steam
32.25 p
23.01 M
Steam
130.9 p
142.4 M
Oxygen
36.92 p
113.4 T
44.09 M
Steam
32.25 p
5.896 M
COS
Hydrolysis
33.32 p
146.7 T
102.3 M
32.75 p
135 T
99.24 M
Cooler 1
Cooler 2
29.99 p
104.5 T
89.67 M
31.1 p
79.44 T
92.14 M
Cooler 3
30.57 p
37.78 T
91.12 M
30.57 p
37.78 T
89.67 M
AGR
Ambient air
1.013 p
15 T
183.3 M
ASU
H2S
1.454 M
Discharge
134.4 M
52.65 M
632.5 T
970.4 M
Fuel
Preparation
Syngas
29.99 p
104.5 T
89.67 M
N2
4.787 M
18.5 p
414.4 T
Ambient air
1.013 p
15 T
880.7 M
17.57 p
1350 T
Stack
1.013 p
151.8 T
970.4 M
HRSG
124 p
566 T
3516 h
169.8 M
1.048 p
633.5 T
485.2 M
387279 kW
286949 kW
ST
GE 7251FB
(Totals of 2 units)
0.0689 p
38.74 T
2440.7 h
198.2 M
p [bar], T [C], h [kJ/kg], M [kg/s], SteamProperties: Thermoflow - STQUIK
GT PRO 16.0.1 Mims
0 09-28-2006 09:46:32 file=C:\Tflow16\MYFILES\SEM04.GTP
Figure 5. Example of a “Shell Process” IGCC with full raw syngas heat recovery
As mentioned previously, there are some concerns about reliability and availability of convective syngas
coolers. If these were dispensed with, and the syngas were just quenched with water, the IGCC efficiency
for this example would drop, drastically, from 45.2% to 40.14%.
20-14
 M. A. Elmasri, 2006-2007
Chapter 20: Introduction to IGCC
20.4.4 “E-GAS PROCESS” EXAMPLE
Figure 6 shows an example of this process. It is very similar to the “GE Process”, except that the gasifier
is of the two-stage type. This is more efficient, producing cooler syngas that possesses a higher calorific
value after the second stage. With full heat recovery from the raw syngas, the example of Figure 6 results
in just about 45% net LHV efficiency, using assumptions very similar to those used in the other examples
shown. Like the others, a drastic drop in efficiency would result from eliminating the raw syngas coolers
and quenching with water instead.
Gasification SystemFlowSchematic - Type 3 Gasifier with Convective Cooler
Pittsburgh No. 8, 4106 tonnes / day
25 T
Fuel
Slurry
47.53 M
Preparation
68.73 M
1.013 p
15 T
169.4 M
ASU
2 Units
Oxygen
36.92 p
113.4 T
40.74 M
15 T
21.2 M
Water
N2
5.171 p
15 T
128.6 M
1015.1 T
35.08 p
1015.1 T
104.3 M
36.92 p
113.4 T
40.74 M
Gross Power
Net Power
Aux. & Losses
Gross Heat Rate
Net Heat Rate
Gross Electric Eff.
Net Electric Eff.
Gasifier Eff.
Fuel LHV Input
Fuel HHV Input
661197 kW
591755 kW
69441 kW
7163 kJ/kWh
8003 kJ/kWh
50.26 %
44.98 %
80.9 %
1315554 kWth
1376258 kWth
ASU: Air Separation Unit
RSC2: Raw Syngas Cooler 2
GCS: Gas Cleanup System
Stage 2
10.31 M
HPE3
130.9 p
326.4 T
1501.9 h
104.2 M
35.08 p
1426.7 T
93.95 M
Stage 1
RSC2 (Convective)
2 Units
58.42 M
33.32 p
371.1 T
104.3 M
GCS
2 Units
29.99 p
166.5 T
90.75 M
To GT
Gasifier
(Type 3)
Totals of 2 Units
HPB1
130.9 p
331.4 T
2664 h
104.2 M
Slag
100 T
5.219 M
p [bar] T [C] h [kJ/kg] M[kg/s], Steam Properties: Thermoflow - STQUIK
0 09-28-2006 11:02:09 file=C:\Tflow16\MYFILES\SEM06.GTP
GT PRO16.0.1 Mims
GCS Inlet
Raw syngas, vol %
CO
39.17%
CO2 11.39%
CH4 0.9858%
H2
33.03%
H2S 0.7831%
O2
0%
H2O 12.99%
COS 0.0343%
N2
1.615%
Ar
0%
GCS Exit
Clean syngas, vol %
CO
45.33%
CO2 13.22%
CH4 1.141%
H2
38.22%
H2S 0.0095%
O2
0%
H2O 0.216%
COS 0.0008%
N2
1.869%
Ar
0%
LHV = 11701 kJ/kg
Figure 6. Example of a “E-Gas Process” IGCC with full raw syngas heat recovery
20.5
AIR-BLOWN IGCC PROCESSES
There are several air-blown IGCC processes in advanced development and testing, with some pilot plants
already in operation. Some of these schemes can, when fully developed, credibly promise efficiencies on
the order of 45% (LHV), coal pile to bus bar,. These efficiencies are similar to what can be realised with
the oxygen-blown systems if they include full heat recovery from the raw syngas. However, it is
important to note that many of the oxygen-blown IGCC projects currently in construction or planning
involve partial or total quench of the syngas with water, instead of heat recovery, and their efficiencies are
thus only in the 37%-41% range.
20-15
 M. A. Elmasri, 2006-2007
Chapter 20: Introduction to IGCC
20.6
COMPARISON WITH CONVENTIONAL PULVERISED COAL
Comparison between IGCC and PC plants
46
45
Net LHV Efficiency, %
44
43
42
41
40
39
38
37
36
IGCC, full syngas heat
recovery
IGCC, radiant
cooler/quench
IGCC, quench
PC, 2RH, adv supercritical
PC, 1RH, supercritical
PC, 1RH subcritical
Figure 7. Efficiency comparison between IGCC and Conventional Pulverised Coal Plants
Figure 7 compares efficiencies between typical oxygen-blown IGCC, with the key variants in raw syngas
heat recovery, and pulverised coal plants. It is assumed that both plant types have sulphur & particulate
capture, but neither has SCR. For the PC plants, sulphur capture is via the wet limestone scrubbing
process, and particulate capture is via an electrostatic precipitator. For the IGCC plants, sulphur and
particulates are captured in the syngas scrubbing and chemical cleanup processes before it is delivered to
the gas turbine.
The parameters of the PC plant models used in this comparison are summarised below. All are proven in
operating PC plants, although the advanced supercritical is not in wide usage.
Plant Model
Key parameters (SI units)
Key parameters (British units)
PC, 2RH, adv supercritical
PC, 1RH, supercritical
280 bar/600°C/600°C/600°C/40 mb
240 bar/570°C/570°C/50 mb
4060 psia/1112°F/1112°F/1112°F/0.58 psia
3480 psia/1048°F/1048°F/0.725 psia
PC, 1RH subcritical
160 bar/550°C/550°C/50 mb
2320 psia/1022°F/1022°F/0.725 psia
It is clear from the above that on an efficiency basis, IGCC based on current technology, but once the raw
syngas convective coolers have been de-bugged, is likely to bring about an improvement over pulverised
coal that is on the order of two percentage points in efficiency over advanced supercritical double reheat
(45% vs 43%). This corresponds to a 6% reduction in coal consumption.. The gain over subcritical
steam plants is about six percentage points (45% vs 39%), representing a 15% reduction in coal
consumption. However, with the syngas quench systems, there is no efficiency advantage for IGCC.
The capital cost of PC plants in the 600-MW range is estimated at about $1800-$2200/kW. The capital
cost for IGCC plants in this size range is estimated to be slightly higher, roughly $2000-$2400/kW.
As IGCC becomes more fully developed, it is expected that their efficiency and reliability will increase,
and their cost will decrease. PC plants, by contrast, are relatively mature although higher material
temperatures and corresponding increases in steam pressures and temperatures are expected to raise their
efficiencies within the next two decades, by perhaps two or three percentage points.
20-16
 M. A. Elmasri, 2008-2009
Chapter 21: Introduction to Carbon Dioxide Capture
INTRODUCTION TO CARBON DIOXIDE CAPTURE
Revised May, 2009
© Maher Elmasri 2008-2009
21.1
GLOBAL WARMING
The earth's atmosphere contains so-called "greenhouse gases". These are gases that allow greater
transmission of energy from the sun to the earth, than radiation from the earth to space. This is because
gases have a different transmissivity for different wavelengths, and the distribution of radiant energy with
wavelength depends on temperature. These gases have greater transmissivity for the wavelengths of the
high-temperature incident solar spectrum, than for the spectrum corresponding to the modest-temperature
radiation from the earth to space. Glass has similar radiation properties and so it is used to construct
greenhouses, so these gases have come to be known as "greenhouse gases" since they cause the earth to
behave like a giant greenhouse.
The presence of greenhouse gases in the atmosphere is natural, and necessary for the earth to be at a stable
temperature that supports the environment as we know it. The problem is that the concentration of these
gases has been rising, causing the earth's temperature to rise potentially to levels that cannot sustain the
present environment.
The principal greenhouse gases are water vapour and carbon dioxide, with water vapour playing the
dominant role in the "greenhouse effect". Secondary greenhouse gases include methane, ozone, and
newcomers like chlorinated fluorocarbons and NOx. Air passing over oceans and lakes picks up water
vapour at a rate that increases with temperature, so as the earth's oceans and lakes become marginally
warmer due to the increase in other greenhouse gases, the concentration of water vapour in the atmosphere
increases, creating a multiplier effect from CO2 , CH4 , etc on global warming.
Fossil fuel combustion is the major contributor to global warming, as it dumps both water vapour and
carbon dioxide into the atmosphere. The amount of water vapour added into the atmosphere directly from
fossil fuel combustion is small compared to the water vapour evaporated from oceans and lakes, but the
effect of the CO2 emitted increases the earth's temperature, thereby increasing the evaporation rates of water
vapour from oceans and lakes.
The concentration of CO2 in the atmosphere has been rising, slowly since the beginning of the industrial
revolution, then at an increasing rate since the era of mass prosperity began, in the USA in the 1950's, in
Europe and Japan in the 1960's, and in the more populous Asian and Latin American countries in the 1980's
and 1990's. Although the efficiency of using fossil fuels has increased dramatically in this modern era,
enabling a greater practical benefit from each unit of fossil fuel consumed, yet the increasing global
population and standard of living has outpaced the improvements in utilisation efficiency, so the
consumption of fossil fuels has been increasing, roughly doubling between 1970 and 2010. The
concentration of CO2 in the atmosphere is believed to have been stable at about 0.03% (300 ppm) for many
millions of years, but the measured values have been rising in recent years, and is now approaching 400
ppm. A considerable body of scientific opinion now believe that this is already causing dangerous global
warming, and that should it reach 450-500 ppm the result will be environmental catastrophe.
As of 2007, global emissions of CO2 are estimated at 27 Gt/yr (27 billion metric tons). The USA is
responsible for about 6 Gt/yr and China is next at about 5 Gt/yr. The EU countries combined emit about 4
Gt/yr. Since China is growing rapidly, it is likely to surpass the USA in this dubious distinction in the near
future.
Figure 1 shows the distribution of CO2 emissions sources in the USA (according to the US DOE). It shows
that power plants are the source of about 40%.
20-1
 M. A. Elmasri, 2008-2009
Chapter 21: Introduction to Carbon Dioxide Capture
Estimated USA CO2 Emissions (2006) - Total 5.9 Gt/yr
Electricity
Generation
9%
17%
Transportation
40%
Industrial
Residential &
Commercial
34%
Figure 1
There is increasing global societal pressure to control CO2 emissions, principally through the notion of
"CCS", an acronym for "carbon capture and sequestration (or storage)".
21.2
CARBON DIOXIDE EMISSIONS FROM POWER PLANTS
The table below summarises the main greenhouse gas emissions from modern, high-efficiency power
plants calculated for two types, a coal-fired steam plant and a natural gas fired combined cycle. The first
two columns show the stack emissions, and the third column shows the water vapour emitted from a wet
cooling tower, if used.
Table 1. Typical values of greenhouse gases produced by power plants
kg/kWh = ton/MWh
Plant Type
Stack CO2
0.8
0.35
Conventional Pulverised Coal
Natural Gas Combined Cycle
Stack H2O
0.25
0.33
CT H2O
1.6
0.8
Both plants are assumed to be utility sized, 500 MW class and operated at their optimum design-point.
Given the losses associated with part-loads, startup/shutdown cycles, and ageing, the actual emissions per
kWh over the plants' life are likely to be 15-20% higher than shown . Since coal properties vary widely, the
figures shown for coal are only indicative.
21.3
CARBON DIOXIDE CAPTURE
The technology to capture CO2 from a mixture of gases has been used for decades in the chemical
industries. However, since the potential application for power plants is new and very large in scale, there
are several new technologies being researched and developed with the goal of making it more efficient and
cost effective.
The basic process is to mix the gases with a liquid solvent that has an affinity to absorb CO2. This mixing is
effected in an "Absorber". A typical absorber is a tower with a falling film of the solvent cascading
20-2
 M. A. Elmasri, 2008-2009
Chapter 21: Introduction to Carbon Dioxide Capture
downwards over a large number of trays in a counter current of the gases. The solvent absorbs CO2 and
leaves the bottom of the absorber in a "CO2-rich" condition. The rich solvent is then pumped to a
"Stripper", where the CO2 is boiled off, either via an increase in temperature or a reduction in pressure, or
both. The stripper is said to "regenerate" the solvent, and the solvent leaves it in a "CO2-lean" condition
before being pumped back to the absorber. Typically, the solvent will need to be cooled and/or
re-pressurised before being returned to the absorber.
There are two basic categories of solvents, chemical and physical, each type being best suited to
summarised below.
21.3.1 CHEMICAL ABSORPTION FROM FLUE GASES
A chemical solvent reacts with the CO2 in the absorber forming an intermediate compound. The rich
solvent is pumped to the stripper, where it is heated by steam to break-up the intermediate compound and
release the CO2, regenerating the solvent. The regenerated solvent, now lean in CO2, is cooled and pumped
back to the absorber.
Vol%
N2 86.31%
O2 5.6%
CO2 1.51%
H2O 5.543%
SO2 0%
Ar
1.038%
CO2CapturePlant FlowDiagram
CWout
25T
15274m
Fluegas out
1.014p
35T
626.5m
15.01T
13658m
25T
13658m
Other
25T
1615.2m
Other
KO
Drum
1.517p
35T
135.7m
CO2: 96.29%
H2O: 3.706%
N2
O2
CO2
H2O
SO2
Ar
Vol%
70.27%
4.559%
12.3%
12.03%
0.0006%
0.8452%
CO2capture: 133.6kg/s, 11546tonne/day
CO2captureefficiency: 90%
Heat input: 466211kW, 466.2MW
Total electrical power consumption: 54229kW
Solvent consumption: 6.234tonne/day
Absorber
45T
CW
Stripper
Steam/CO2
Richsolvent
2004.4m
Reboiler
Water fromabsorber
and/or knockout drum
35.06m
Ps = 55.25bar
Ts = 287.4C
3.809p
Solvent pumps: 3102kW
CO2
151.7p
35.01T
133.6m
Vol%
CO2: 100%
H2O: 0%
Drain
2.085m
Condensate
Gas
Cooler
CWin
15T
15274m
1615.2m
CO2Compressor
46298kW
Makeup
solvent
0.0722m
Fluegas in
1.139p
52.13T
797.3m
3123kW
35T
37.15m
Steam
3.809p
220.5T
2904.1h
202.2m
Condensate
15.54p
142.1T
598.8h
202.2m
380.2kW
p[bar] T[C] m[kg/s] h[kJ/kg]
STEAM PRO 19.0 Maher Thermoflow, Inc.
0 04-28-2009 15:12:31 C:\Sem_9-08\Sec21_Solar\Generic 500 MWnet w all defaults CO2 capture.stp
Figure 2. Schematic of an MEA process to capture CO2 from a 500 MW coal-fired conventional plant
The strength of the chemical bond between the solvent and the CO2 makes it feasible to absorb dilute CO2 ,
such as from the flue gases of a power plant at atmospheric pressure. However, this strong bond also
means that more thermal energy is needed to breakup the intermediate compound and separate the CO2
from the solvent in the stripper. This energy is typically in the form of steam, which must be extracted from
the power plant cycle, thereby reducing its output and efficiency.
The most proven chemical solvent is mono-ethanol-amine (MEA). In practice, this process requires a heat
input about 1.5 kg(lb) of steam at 3-5 bar (45-75 psia) per kg(lb) of CO2 captured. This required heat input
is a function of the CO2 capture efficiency as well as the initial CO2 concentration in the flue gases.
In a typical coal fired power plant, the MEA process will need a steam flow rate that represents roughly
35% of the HPT inlet flow, and about 65% of the steam mass flow rate exhausting to the condenser of the
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 M. A. Elmasri, 2008-2009
Chapter 21: Introduction to Carbon Dioxide Capture
original plant (without MEA). Thus, if the steam flow needed for the stripper were extracted from the LPT,
this enormous reduction in condensing flow may require a re-design of existing steam turbines.
Other chemical solvents have been used and are in various stages of development with the objective of
reducing the energy required in the stripper and produce better overall economics. A system using chilled
ammonia as the solvent appears to promise a considerable reduction in the consumption of steam at the
stripper, but is not yet as proven as MEA.
Since MEA is the most proven method, we need to recognise that its use implies a very significant reduction
in steam plant performance, as well as increase in cost. Calculations show that adding MEA to an existing
conventional coal-fired plant will reduce its net output and efficiency by roughly 25%. This means that in
order to get the same final net output into the grid, coal consumption increases by 33%. It also means that
the capital costs shall increase by the cost of all the new equipment needed for CO2 capture, plus another
33% to buy a larger power plant that, after losses, ultimately feeds the same power to the grid.
Table 2 compares key parameters of a hypothetical, green-field conventional coal power plant, designed
with 90%-efficient MEA CO2 capture, relative to a similar plant designed without CO2 capture. In this
comparison, both plants are designed to deliver the same net power to the grid, 500 MW after all auxiliaries.
Table 2. Coal-fired plant with CO2 capture by MEA relative to a plant without. Both provide same net
output to the grid. η = 42% (LHV), reduced to η ≈ 32% (LHV) with CO2 capture.
Net Power Output (500 MW)
same
Coal Consumption (and net heat rate)
+34%
Steam Turbine Output
+14%
Steam Turbine Inlet Mass Flow rate
+34%
Steam Turbine Exhaust Mass Flow rate
-40%
Total Cooling Water Heat Rejection
+90%
Estimated Capital Cost per net kW
+70%
Final CO2 emitted per net MW
-87%
Table 3 compares a hypothetical, green-field natural gas fired combined cycle, designed with 90%-efficient
MEA CO2 capture, relative to a similar plant designed without CO2 capture. In this comparison, both plants
are designed with identical gas turbines, so their net power outputs are dissimilar.
Table 3. Natural gas fired combined cycle with CO2 capture by MEA relative to one without. Both have
identical gas turbines. η = 56.2% (LHV), reduced to η ≈48.5% (LHV) with CO2 capture.
Net Power Output
-14%
Natural gas input
same
Steam Turbine Output
-20%
Steam Turbine Inlet Mass Flow rate
same
Steam Turbine Exhaust Mass Flow rate
-44%
Total Cooling Water Heat Rejection
+54%
Estimated Capital Cost per net kW
+109%
Final CO2 emitted per net MW
20-4
-83%
 M. A. Elmasri, 2008-2009
Chapter 21: Introduction to Carbon Dioxide Capture
As the above tables show, using the proven MEA process for CO2 capture is expected to incur steep
penalties to performance and cost. Thus, this author is sceptical about its viability, and believes that
post-combustion CO2 capture from power plants will only become viable if superior methods are
developed, such as chilled ammonia, or others, that can significantly lower these penalties.
21.3.2 PHYSICAL ABSORPTION
A physical solvent simply dissolves the CO2, much like water does, except that it can hold greater
concentrations of CO2 in solution. Solubility of gases in liquids increases with pressure, so to use the
typical solvent, the gas stream containing the CO2 is pressurised then mixed with solvent in the absorber.
The solvent pressure is reduced in the stripper, causing some (but not all) of the dissolved CO2 to come out
of solution. This is similar in principal to what happens when a can of soda is opened, reducing the pressure
and allowing CO2 to escape from the solution.
The physical solvent process works best when CO2 is in high concentrations in the gas stream, and when the
entire gas stream is at high pressure. This makes it unsuitable for capturing CO2 from the stack gases of a
typical power plant, but makes it well suited to pre-combustion CO2 capture from the syngas in an IGCC.
In a typical IGCC scheme with CO2 capture, the high-pressure syngas is put through a "water-gas" shift
reaction, converting the CO into CO2 and hydrogen via the exothermic reaction:
CO + H2O → CO2 + H2 + heat
For this reaction to occur effectively, a catalyst is used, and the feedstocks entering the reactor are at about
260-370˚C (500-700˚F). Although the reaction combines one mole of H2O with one mole of CO, in
practice, excess H2O allows a greater conversion ratio, so a H2O to CO ratio of 2:1 (or greater) is typical.
This requires moisturising the syngas before the shift reactor, such as by injecting steam. In IGCC schemes
where water quenching is used instead of a convective syngas cooler, the required excess H2O is achieved
with less steam injection. Mixing steam with the syngas prior to the shift reactor lowers system efficiency,
since the steam could otherwise have produced work in the power cycle. System efficiency is also reduced
by the fact that the hydrogen syngas produced and sent to the gas turbine has a lower calorific value than the
CO consumed in the reaction. Since the reaction is exothermic, this energy difference appears as heat
which may be recovered, but its use in the cycle as low-grade heat is less efficient than having the same
energy reach the gas turbine where it is used as high grade heat.
Following the shift reaction, the high pressure syngas, typically at around 40 bars (600 psia) is scrubbed
with the solvent. Typical solvents include "Selexol" (di-methyl-ether of polyethylene glycol) and
"Rectisol" (cold methanol), as well as others. The solvent is then reduced in pressure to release the CO2 in
the stripper. To get most of the CO2 out of solution, requires a substantial pressure reduction, typically to
about 3-4 bars (45-60 psia). However, since the CO2 will need to ultimately be compressed to about 150
bars (2250 psia) to send it into a pipeline, it is wasteful to reduce the pressure in one stage, so several stages
of pressure reduction are employed, with some of the CO2 flashing off at each pressure. A practical system
may thus have 3-5 stages of pressure reduction with only about one-third of the total CO2 emerging at the
lowest pressure stage.
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 M. A. Elmasri, 2008-2009
Chapter 21: Introduction to Carbon Dioxide Capture
AcidGas Removal andCO2Capture- Eachof 2Units
Vol%
CO 1.111%
CO2 5.678%
CH4 0.0034%
H2 90.74%
H2S 0.0109%
O2 0%
H2O 0.0458%
COS 0.0011%
N2 1.845%
Ar
0.5652%
M.W.= 5.395
CWout
25T
1028.1m
Syngas out
23.37p
35T
12.38m
Vol%
CO 0.6797%
CO2 41.37%
CH4 0.0021%
H2 55.52%
H2S 0.6691%
O2 0%
H2O 0.2801%
COS 0.0007%
N2 1.129%
Ar
0.3458%
M.W.= 20.25
Syngas in
23.37p
37.78T
75.94m
8.217p
24.58m
Flash
Tank3
CO2
Absorber
Other
0.7251p
21.51m
Stripper
LeanSolvent
H2Sremoval auxiliary: 1679.5kW
CO2
151.7p
35.01T
61.45m
CO2Compressor
17087kW
Acidgas
1.52p
50T
2.019m
Vol%
H2S: 45.94%
CO2: 45.94%
H2O: 8.116%
M.W.= 37.34
CW
KO
Drum
CO2-RichSolvent
H2S/CO2Rich
Solvent
CO2capture: 61.45kg/s, 5309tonne/day
CO2captureefficiency: 90%
H2Sremoval: 0.8465kg/s, 73.14tonne/day
H2Sremoval efficiency: 99%
Heat input: 17719kW, 17.72MW
Total electrical power consumption: 25074kW
4.834p
15.36m
Flash
Tank1
H2S
Absorber
210.2kW
15.01T
Other
Flash
Tank2
Semi-Lean
Solvent
CWin
15T
1028.1m
Condensate
Steam/ Acidgas
Reboiler
Ps = 48.95bar
Ts = 279.7C
3.375p
0.6406kW
Steam
3.375p
228.3T
2922.3h
7.561m
Condensate
3.906p
137.6T
579h
7.561m
p[bar] T[C] m[kg/s] h[kJ/kg]
GT PRO 19.0 Maher
0 05-01-2009 15:58:23 file=C:\Sem_9-08\Sec21_Solar\GTCC 2x7FB with IGCC-Selexol.gtp
Figure 3. Schematic of a Selexol process to capture CO2 from syngas of an IGCC
After scrubbing, the syngas chemical energy is largely in the form of hydrogen, and this is burnt in the gas
turbine with minimal production and release of CO2 to the environment.
IGCC with CO2 capture is still a developing technology. Table 4 below presents an estimate for the output
and efficiency losses incurred with CO2 capture. In this comparison, the IGCC without CO2 capture is
assumed to have both radiant and convective syngas coolers for maximum efficiency (44.6% LHV). The
IGCC with CO2 capture, however, has only a radiant syngas cooler, and the convective syngas cooler is
replaced by water quenching to increase the H2O to CO ratio prior to the shift reaction, reducing the
necessary steam addition at the shift reactor itself.
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 M. A. Elmasri, 2008-2009
Chapter 21: Introduction to Carbon Dioxide Capture
Table 4. IGCC with CO shift reactor and Selexol CO2 capture( η ≈ 36.7% LHV) relative to IGCC with
Radiant Syngas Cooler and no carbon capture ( η ≈ 44.6% LHV), both based on same GT model.
Net Power Output
-11%
Coal Consumption
+8%
Coal Consumption per net kWh (net heat rate)
+21%
Estimated Capital Cost per net kW
+20%
Final CO2 emitted per net MW
-88%
21.3.3 OTHER METHODS OF CO2 CAPTURE
There are other approaches to capture CO2 from a gas stream. They include adsorption into a bed of solids,
membrane systems, etc, but the purpose here is to focus on the technologies which are considered most
likely on a commercial basis in the near term.
21.3.4 OXYFUEL CYCLE
Instead of separating the CO2 from the flue gas stream, this approach calls for making a flue gas stream that
is nearly all- CO2. An air-separation plant produces oxygen which is used to burn the coal in a conventional
coal plant. The flue gas is then mostly CO2, along with water vapour from the hydrogen in the coal, as well
as some excess oxygen. Additionally, there will be some other gases arising from the coal, like nitrogen
and sulphur oxides. Since the oxygen from the air separation plant will still have some nitrogen and argon,
these will also be present in the flue gases.
After the flue gas leaving the boiler is de-sulphurised, such as by a wet limestone scrubber, and the ash
particles removed, it is then split, with roughly 80% being recirculated to the boiler inlet where it is mixed
with the oxygen from the air separation plant and fed into the boiler for combustion. Thus, what enters the
boiler resembles air in its oxygen content, except that CO2 instead of N2 is the main diluent. The 20% of the
flue gas that is not recirculated to the boiler inlet is mostly (say 90%) CO2 and this can be compressed and
disposed of into the CO2 disposal pipeline. The non-CO2 components can be separated out during the
compression process which takes place in multi-stages with intercoolers. Water will condense and be
knocked out at modest pressures, and at higher pressures, CO2 condenses or becomes supercritical with
much higher density than nitrogen and oxygen, so it can be separated by gravity in flash tanks.
Alternatively, chemical scrubbing or a solvent processes can be used to separate the CO2 from the O2 , N2
and Ar.
This technology is work-in-progress. The table below represents the author's estimate for the output and
efficiency losses incurred when a conventional coal plant is reconfigured as an oxy-fuel plant for CO2
capture. It suggests that oxy-fuel systems are likely to be more efficient than applying MEA scrubbing to
the flue gases of a conventional coal plant.
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 M. A. Elmasri, 2008-2009
Chapter 21: Introduction to Carbon Dioxide Capture
Table 5. Estimated reductions in plant output and efficiency due to CO2 capture by reconfiguring a
conventional coal plant as an oxy-fuel system. Base η = 42% (LHV), with CO2 capture η ≈ 34% (LHV)
Air separation plant
Misc auxiliaries
Final CO2 compression for pipeline
Total losses due to CO2 capture
21.4
8-10%
about 3%
6-8%
17-21%
CARBON DIOXIDE SEQUESTRATION
Once the CO2 is captured, the next phase is to dispose of it in an ecologically sound fashion.
Many geological formations include gases trapped within the pores of the rocks and soil. Natural gas is an
obvious example, but there are also naturally occurring CO2 gas fields, some of which have been tapped and
utilised. In these geological formations gases have been trapped within the pores for very long time scales.
Thus, if CO2 were injected into such formations, it is believed that it will be retained for geological time
scales. Additionally, CO2 can be injected into saline aquifers, where it will dissolve and be sequestered.
Some geologists believe that there are sufficient formations suitable for capturing and retaining all CO2
emissions that result from burning all the fossil fuels between now and the end of the fossil fuel era.
Due to the high cost and efficiency penalties, the present author doubts that CCS on every significant power
plant in the world is going to be economically acceptable. However, injecting CO2 into the ground to
displace oil and/or hydrocarbon gases that otherwise are not economically recoverable seems like a very
viable proposition. So-called "Enhanced Oil Recovery" or EOR is a technique to inject hot water, steam or
CO2 into hydrocarbon fields to drive trapped hydrocarbons out of depleted fields or out of geological
formations from which the hydrocarbons cannot otherwise flow freely. In such cases, if CO2 is used as the
EOR fluid, it can displace the hydrocarbons and be retained within the geological formations after the
hydrocarbons have been driven out. In this context, CO2 sequestration seems quite likely to succeed,
eventually on a large scale.
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 M. A. Elmasri, 2008
Chapter 22: Introduction to Solar Thermal
INTRODUCTION TO SOLAR THERMAL ENERGY
October, 2008
© Maher Elmasri 2008
22.1
ABUNDANCE OF SOLAR ENERGY
The solar energy flux reaching the outer edge of the earth’s atmosphere is about 1.37 kW/m2. The
atmosphere absorbs some of this energy, and the rest reaches the surface. On a clear day at solar noon,
when the sun is directly perpendicular to the earth’s surface, the energy flux reaching the surface is about 1
kW/m2. When the sun is not directly perpendicular to the earth’s surface, less energy is received per square
meter of horizontal surface because (a) the path of the solar rays through the atmosphere is longer, and (b)
the projected area of a horizontal square meter facing the sun is diminished.
Solar energy is quite abundant. The projected area of the earth facing the sun is (π/4)Dearth2, and the
diameter of the earth is 12,760 km; so at a flux of 1 kW/m2, we can estimate that the earth receives about
1.28 x1014 kW of solar energy. There are 8760 hours in a year, so we find that the earth receives 1.12 x1018
kWh of solar energy per year. For 2007, all human consumption of primary energy from all sources is
estimated at 1.4 x1014 kWh, i.e. about 0.0125% of the solar energy reaching the earth. Another way of
looking at this number is that human raw energy consumption in a year is on the same order of magnitude as
solar energy reaching the earth in an hour.
Because of the diffuse nature of solar energy, harnessing it is complex and expensive. Broadly speaking,
there are three methods for humans to harness this energy, as summarised below.
22.1.1 BIOMASS
This is the oldest and most natural method. Vegetation stores solar energy by creating hydrocarbons from
water and carbon dioxide. If the vegetation is left to naturally decay, it will eventually release this chemical
energy to the ambient as heat, as the hydrocarbons are slowly and naturally oxidised and the resulting water
and carbon dioxide are returned to the environment. However, if humans harvest the vegetation, they may
oxidise it (burn it) in a power plant and use the resulting energy to their benefit before it is re-released to the
environment.
22.1.2 PHOTOVOLTAICS
Solar panels containing a vast number of photo-diodes are used to directly convert solar energy to
direct-current electricity. In simple terms, the photons impinging on a semiconductor material increase the
energy of the electrons above a certain threshold, enabling them to flow against a potential. Only a portion
of the solar energy can be converted to electricity in this fashion, and the rest appears as heat. This is due to
the fact that not all electrons are energized to the requisite threshold level, so their energy is dissipated; as
well as due to the internal resistance within the material. Various materials are used to construct solar cells.
Most of the mass-produced and less expensive cells are based on silicon, and they convert solar energy to
electricity at efficiencies in the range of 10-16%. More advanced multi-junction cells in the laboratory and
pilot project stage can achieve conversion efficiencies of up to 35%.
The trade group EPIA (European Photovoltaics Industry Association) reports Global installed PV capacity
as 9,000 MW (peak) as of the end of 2007. This is up by a factor of ten, from under 1,000 MW (peak) in
1997. They also report that about half the global capacity is within the EU and have stated a goal of
supplying up to 12% of Europe’s electricity from photovoltaics by 2020. The USA has lagged behind in
this area, and the DOE EIA reports photovoltaic installed capacity was only 340 MW as of the end of 2006,
but this also has been increasing very rapidly.
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 M. A. Elmasri, 2008
Chapter 22: Introduction to Solar Thermal
22.1.3 SOLAR THERMAL
Using solar energy as a source of heat is quite basic. Collecting it for low-grade heat applications, such as
domestic hot water or to heat buildings is very simple. A basic system that uses water (or ethylene glycol)
is just a coil of tubes thermally bonded to a backing plate and painted a high-absorptivity black. The coils
are installed on roofs that face the sun, and are normally placed behind a glass plate to reduce convective
heat losses from the wind. The coil enclosure may be evacuated to further limit heat losses to achieve
higher temperatures and efficiencies, but naturally this increases the cost. Other systems use air as the heat
transfer medium for houses.
Using solar heat for power generation requires higher temperatures, which necessitates concentrating the
solar energy using mirrors to focus heat collected over a large area onto a smaller receiver, as further
discussed below.
22.2
SOLAR THERMAL POWER GENERATION
22.2.1 TYPES OF SOLAR CONCENTRATORS
Broadly speaking, there are two categories of solar concentrators, those that focus on a “point” and those
that focus on a “line”. Point-concentrators focus thermal energy onto a small area, enabling very high
temperatures to be achieved. An example of this sort of concentrator is shown in Figure 21.1.
Figure 1. Example of a point concentrator, in this case a high-temperature solar furnace at Odeillo, France, capable
of achieving extremely high temperatures (Photo in the public domain, after Wikipedia under the GNU Free
Documentation License)
Other geometries of point-concentrators have been used and/or proposed for power generation, with
mirrors on the ground focusing heat onto a receiver atop a tower. The receiver may be a boiler to power a
steam cycle, and in such designs steam conditions are comparable to those used in power plants employing
conventional fired boilers (steam temperatures of 500-600 ˚C). Alternatively, the receiver may be an air
heater to power a gas turbine, but in such designs it is difficult to obtain peak temperatures similar to fired
gas turbines (1250-1450˚C), since heat at such temperatures cannot be transferred across nickel-based alloy
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 M. A. Elmasri, 2008
Chapter 22: Introduction to Solar Thermal
tubes, and ceramics must be used. A small solar gas turbine pilot plant has been built with a porous ceramic
receiver.
Line concentrators focus the energy onto a line. There are two geometries of line concentrators in use and
development: Parabolic Trough concentrators and Fresnel Concentrators.
Figure 2. Example of a Parabolic Trough Concentrator (Courtesy of Solel)
Parabolic Trough concentrators are shown In Figure 2, and were used in the first large scale solar power
plant, built in the mid-1980’s at Kramer Junction (Figure 3). The parabolic mirrors focus the energy on a
pipe that carries thermal oil, and the pipe is encased in an evacuated glass tube to minimise convective heat
loss from the hot oil. The hot oil is used to generate steam to power a turbine.
The advantage of the parabolic trough system is its high efficiency and high geometric concentration. Its
principal disadvantage is its cost and the complexity of the sun tracking system as the troughs are large and
unwieldy.
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 M. A. Elmasri, 2008
Chapter 22: Introduction to Solar Thermal
Figure 3. 44-MW solar power plant using parabolic trough concentrator (Courtesy of Solel)
The second type of line concentrator is the Frsenel system, which consists of a larger number of smaller flat
mirrors to focus the energy onto the receiving tube, as shown in Figure 4. The primary advantage of this
system is simplicity and moderate cost. Another advantage is that the receiving tube is mounted separately
from the mirrors, making it easier to generate steam directly within the receiving tube without the
intermediation of a thermal oil. However it is less efficient and provides a lower geometric concentration
than the parabolic trough system.
Figure 4. Fresnel
Concentrator
(Courtesy of SPG)
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 M. A. Elmasri, 2008
Chapter 22: Introduction to Solar Thermal
Figure 5. Fresnel Mirror Focusing Controls (Courtesy of Novatec-Biosol)
Both types of linear concentrators can provide heat at temperatures consistent with typical steam power
plants. However, most thermal oils available are limited to under 400˚C, so if steam temperatures above
roughly 370˚C are to be achieved, then a higher-temperature heat transfer fluid or direct steam generation
within the receiver tubes must be used. Direct steam generation within the receiver tubes increases pressure
drops and poses risks of leakage. Molten saltpetre (a salt that is a mixture of Sodium Nitrate and Potassium
Nitrate) can be used as a heat transfer fluid for temperatures above 300˚C.
22.2.2 EFFICIENCY OF SOLAR THERMAL STAND-ALONE PLANTS
Table 1. Estimated Efficiencies of solar thermal power generation
Plant Type
Stand-Alone
Steam Cycles
"Solar
Combustor"
GTCC
Typical
Unit Size
Max Working
Fluid
Temperature,
˚C
Net Efficiency,
Collected Heat
to Electric
Power
Overall Net
Efficiency,
Solar Energy
to Electric
Power
25 MW
365
32%
24%
50 bar/365 C
40 MW
480
35%
26%
60 bar/480 C
100 MW
540
40%
30%
100 bar/540 C/28 bar/540 C
500 MW
580
45%
34%
280 bar/580 C/68 bar/600 C
35 MW
830
40%
30%
PR=11, possible metal HX
100 MW
1100
48%
36%
PR=12 (2P NRH)
500 MW
1350
55%
41%
PR=16 (3P RH)
Notes
The table shows estimated efficiencies for various technologies. In calculating cycle efficiencies shown in
the fourth column, ambient is assumed to be 25˚C, sea level, 60% humid, and the steam condensing system
is assumed to employ a wet mechanical draft cooling tower. If an open loop sea-cooled condenser were
used, efficiencies would be higher, and if a dry air-cooled condenser were employed efficiencies would be
lower. In calculating overall efficiencies shown in the fifth column, it is assumed that collection efficiency
is 75% (i.e. ¾ of the solar energy on the collector aperture is transferred to the working fluid).
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 M. A. Elmasri, 2008
Chapter 22: Introduction to Solar Thermal
22.2.2.1 Stand-Alone Steam Cycle with Thermal Oil
Hitherto, the most proven way to collect thermal energy with line concentrators is via thermal oil passing
through the line-collector receiver tubes, with the hot oil generating steam in a boiler. This limits steam
temperature to about 365˚C, which works reasonably with modest-sized units in the 15-50 MW size range
and with steam pressures in the 30-50 bar range. This sort of system is represented by the first row in Table
1. Steam cycle internal efficiency is about 32%. With 75% collector efficiency this corresponds to about
24% overall efficiency, about double the efficiency of proven photovoltaic systems.
22.2.2.2 Stand-Alone Steam Cycles with Higher Temperatures
The next three steam cycles in Table 1 employ temperatures higher than feasible with currently available
thermal oils. Thus, they require either direct steam generation in the receiver tubes or a high-temperature
molten salt.
22.2.2.3 “Solar Combustor Gas Turbine” Combined Cycles
The efficiency of converting heat to power depends on the temperature of the working fluid. Although solar
heat can be concentrated to provide very high temperatures via point-concentrators, the cost and complexity
of utilising the heat increase with temperature.
The first case shown in this category in Table 1 uses a gas turbine with a solar air heater at a modest 830˚C
turbine inlet temperature, similar to gas turbine practice of the 1950’s with uncooled blading. The author
believes that at this modest temperature, the concentrated solar heat exchanger replacing the combustor
may be constructed using Nickel-based alloys.
The second and third cases in Table 1 uses gas turbines with solar air heaters that can produce turbine inlet
temperatures more in-line with modern fired gas turbines. These temperatures require ceramic heat
exchangers that have yet to be developed and proven, although some innovative concepts have been tested
in research plants.
22.2.3 DEALING WITH SOLAR PERIODICITY IN SOLAR THERMAL STAND-ALONE PLANTS
Since the solar heat source is not constant, a stand-alone solar plant will either be an unreliable element on
the grid or will need to have means of extending its operation. The two obvious methods are (a) Thermal
storage, and (b) Provision of a supplementary fuel-fired heat source. Both add cost and complexity.
If thermal oil or molten salt is used as a collection medium, then it can be stored in insulated tanks or
underground caverns
With a steam cycle efficiency of 33%, the amount of heat needed to be stored to generate 100 MWh of
electricity is about 300 MWh of heat, i.e. 300,000 x 3600 kJ = 1.08 x 109 kJ. Thermal oil has a specific heat
of about 2.2 kJ/kg-˚C, so to store this amount of heat through a temperature range of 60˚C, we need to store
about 8,200 tons of thermal oil, which corresponds to about 11,000 cubic meters. The corresponding
amounts with molten salt are about 12,000 tons and 6,500 cubic meters. 100 MWh of electricity
corresponds to just 10 MW for 10 hours, i.e. overnight storage for a 10 MWe plant, or an hour's worth of
energy for a 100 MWe plant.
These vast quantities of storage are problematic. Thermal oils are expensive; and molten saltpetre is
corrosive in the presence of water. Additionally nitrates, being oxidants, are one component of basic
explosives, which are just a mixture of an oxidant and fuel.
It has also been proposed to store heat as high pressure steam or via a high-temperature phase-change
medium, but no matter which method is used, thermal storage is an expensive proposition.
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 M. A. Elmasri, 2008
Chapter 22: Introduction to Solar Thermal
The alternative to thermal storage is to equip the plant with a supplementary boiler that replaces the solar
source as needed. While this is less "green", the author believes it to be more practical, especially in the
early years of solar power.
22.2.4 EFFICIENCY OF SOLAR THERMAL ADD-ON TO EXISTING PLANTS
An excellent way to use solar thermal power is to integrate it with existing thermal power plants. In this
way, whenever solar heat is available it will offset some of the fuel consumed, and whenever it is
unavailable, the plant can operate normally.
This approach eliminates the need for energy storage and reduces the cost of the plant relative to its output.
To maximise economic benefits, the solar contribution should be modest relative to the existing thermal
plant, say 5-25% of its thermal input. This allows the solar field to be utilised to its maximum capacity at all
times without significantly affecting the plant's normal mass and energy balances. Additionally, since the
solar field utilses existing power generation equipment and associated infrastructure, the cost and
complexity are reduced. The major drawback, however, is the need for large areas of land near existing
thermal power plants.
Table 2. Estimated Efficiencies of solar thermal integration with existing plants
Method of Solar Integration with
Existing Plant
Parallel generation of saturated HP
steam in a 3PRHCC
Parallel by-pass heating of HP FW in
conventional steam plant
Parallel steam generation in
conventional steam plant
22.3
Typical Unit
Size
Max Working
Fluid
Temperature,
˚C
Net Efficiency,
Collected Heat to
Electric Power
Overall Net
Efficiency,
Solar Energy to
Electric Power
300 MW
330
45%
34%
500 MW
300
42%
32%
500 MW
420
47%
35%
EFFECT OF LATITUDE & SEASON ON SOLAR HEAT
Most human population and energy consumption occurs in the latitudes between roughly 30˚N and 50˚N,
but the most effective geographic zone for solar energy is between the tropics, ±23.5˚. As one moves away
from this zone, the solar flux weakens. Figure 6 gives an estimate of the energy that can be collected per
square meter of aperture as a function of latitude. The curve labelled "average" is time-averaged over the
entire year, day and night, and with the assumption of modest haze; whereas the curve labelled "peak"
represents the maximum at noon on a clear day. As the diagram shows, the time-averaged heat collection is
only 25%-30% of the peak, so a stand-alone plant without storage would have an unsatisfactory capacity
factor for its power generation equipment.
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 M. A. Elmasri, 2008
Chapter 22: Introduction to Solar Thermal
Estimated solar heat collection per sq.m. of aperture with 75%-efficient
horizontal parabolic trough collectors oriented N-S and tracking E-W
700
Heat Collected, W/sq.m.
600
500
Peak
400
Average
300
200
100
0
0
10
20
30
40
50
Latitude, degrees
Figure 6
Ratios of seasonal averages to the annual average heat collection rate with
horizontal collectors oriented N-S and tracking E-W
160.0%
140.0%
120.0%
100.0%
Summer
80.0%
Spring & Fall
Winter
60.0%
40.0%
20.0%
0
10
20
30
40
50
Latitude, degrees
Figure 7
Figure 7 shows how the seasonal averages for heat collected deviate from the annual average as a function
of latitude. This illustrates the capacity problem as one moves away from the tropics, for example, at 40˚N,
20-8
 M. A. Elmasri, 2008
Chapter 22: Introduction to Solar Thermal
the winter-averaged heat collected is only 30% of the annual average, whereas the summer-averaged value
is 145% of the annual average, so the heat collected in winter is only 1/5th of the heat collected in summer.
Solar thermal technology is still in its infancy, so accurate cost estimates are difficult to predict. However,
as a rough guide, Figure 8 indicates the approximate cost of solar heat. The sun, of course, is free, so the
cost of the heat is that of the collection system. As of 2008, the cost of a linear collection system was on the
order of €300/square meter including a basic thermal oil system without significant storage. Assuming that
the annual cost is 10% of this amount, i.e. €30/annum per square meter, and assuming no cost for land, one
obtains Figure 8. It suggests that the cost of "solar fuel" is of the same order of magnitude as natural gas.
However, the novelty of the technology, the need for large areas of land, and the variable nature of the heat
source all combine to favor natural gas until the environmental issues or the cost of fuel compel us to use the
sun instead.
Estimated cost of solar "fuel" based on 30 €/yr/per sq.meter (300 €/sq.m of
aperture, simple ten-year amortization, no land or operating costs)
14
Estimated Cost of Solar Heat
12
10
$/MMBTU
8
€/GJ
6
4
2
0
0
10
20
30
40
50
Latitude, degrees
Figure 8
For a parabolic trough collection system, the aperture area is about 1/3rd of the land area covered. With an
annual average heat collection of 180 W/sq.m of aperture, we find that one square km of land can support
enough collectors to produce about 66 MW of useful thermal energy, averaged over the entire year. This
corresponds to an annual average electricity generation of about 20-30 MW.
One may then estimate that the entire world's electricity production for 2007 can be generated from an area
of about 100,000 square kilometers (320 km x 320 km or 200 miles x 200 miles). The entire world's total
primary energy supply can be collected from an area of about 270,000 square kilometers. The Sahara desert
covers about 9 million square km, so about 1% of its area would suffice to generate the world's electricity,
and 3% of its area would suffice to collect the total global consumption of primary energy.
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