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A SEMICLOSED CYCLE GAS TURBINE WITH
CARBON DIOXIDE- ARGON AS WORKING FLUID
111 WWI Hlli
Inaki Ulizar
Industrie de Turbopropulsores - Ajalvir
Torrejon de Ardoz
Madrid, Spain
Pericles Pilidis
School of Mechanical Engineering
Cranfield University
Cranfield, Bedford, U.K.
ABSTRACT
This paper describes the performance analysis of a
semi closed cycle gas turbine. The working fluid is carbon
dioxide and the fuel is low heating value gas synthesised
from coal. The objective of the machine is to produce clean
electricity with the smallest efficiency penalty..
Firstly the thermodynamic properties of the gases in
the cycle Were obtained as a function of temperature and
pressure. Then two performance simulation codes were
developed. These have the ability of simulating different
configurations of open, closed and semi-closed cycles. The
first code was used for cycle optimisation and the second for
off-design studies.
The design and off-design performance of the
machine are predicted. The production of clean electricity
will be at the expense of a lower efficiency compared with
current equipment. Finally, some critical issues for the
development of such a gas turbine are identified
INTRODUCTION
The continuing concern over the emission of
greenhouse gases coupled with the ever increasing demand
for electrical energy poses a very difficult challenge to power
engineers. One possible solution to these conflicting
requirements, if solid fossil fuels are to be employed, is to
collect and dispose of the emissions in a controlled way.
For example an internal combustion semi-closed
power cycle, where the working fluid is carbon dioxide, with
oxygen and fuel injected in the combustion chamber and
excess carbon dioxide collected at the outlet seems to be a
very attractive environmental proposition. Such a cycle has a
dual advantage. By collecting and storing safely the excess
carbon dioxide, the emission of greenhouse gases is
controlled. By having a clean fuel and no air in the working
fluid, there are no emissions of oxides of sulphur and
nitrogen.
This proposition becomes particularly attractive
when coal is considered. This fuel is in plentiful supply and
when gasified it can be a very clean source of energy. The
major drawbacks of such a cycle are the complexity of the
equipment and the significant reduction in efficiency caused
by the need to produce pure oxygen.
THE ENGINE
This paper describes an internal combustion turbine
where the working fluid is carbon dioxide, the main mass of
which is recirculated in the engine. The choice of working
fluid is dictated by the fuel based on coal, and the
requirement to make carbon dioxide collection an easy
process.
Heat addition is achieved by injecting oxygen and
coal gas fuel in the combustion chamber. To keep a constant
mass flow through the device water is collected at the outlet
of the cooler and some carbon dioxide is extracted at the HPC
delivery. This gas is compressed to at least 60 atmospheres
for efficient collection. All calculations include this
compression work.
Figure 1 shows the engine selected and described in
the paper, a two-spool gas generator with an independent
povrer turbine. It was selected between several alternatives
with core cycles, incorporating a variety of additional
components, such as intercoolers and heat exchangers.
Presented at the International Gas Turbine and Aentengine Congress & Exhibition
Birmingham, UK — June 10-13, 1996
This paper has been accepted for publication in the Transactions of the ASME
Discussion of it will be accepted at ASME Headquarters until September 30, 1996
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The working fluid includes, in the turbine, any
water produced in the combustion process. A small amount
of Argon is also present because it is much cheaper to leave
this gas in the oxygen stream fed to the combustor than to
separate it. Its presence has a minor effect on the
performance of the cycle.
The gas properties required are enthalpy, entropy
and constant pressure specific heat. Usually they are
correlated as a function of temperature, while the effect of
pressure is assumed to be negligible. However, this
assumption is not valid in high pressure and low temperature
cases. This situation could arise in the cycle under
investigation, so real gas analysis was employed where both
pressure and temperature were used as correlation
parameters. For example Figure 2 shows the Specific Heat at
Constant Pressure (Cp) of carbon dioxide and Figure 3 when
does it behave as an ideal or real gas.
Fuel
HPT
4—
CO2+Ar
Cooler
LPT
PT
Boiler
CO2+Ar H20
Figure 1. Diagram of the Semi-Closed Cycle Engine
GAS TURBINE OVERVIEW
Two development options are available. The first
one is to use existing equipment, the second is to design a
new engine. The first option will be cheaper to develop
while the second will have better performance.
To select the main cycle parameters it was
necessary to have an overview of gas turbine current practice.
To achieve this, information was obtained on more than 500
PROPERTIES OF THE GASES IN THE CYCLE
different models of open cycle gas turbines and combined
cycles, from over 40 manufacturers. The efficiency and
power output of these gas turbines, without bottoming cycles,
are shown in figure 4. Several observations can be made.
Most gas turbines are of the simple cycle type and
only a few are recuperative or intercooled. A low cost option
The thermodynamic properties of the gases in this
cycle, as well as others that could be introduced to enhance
the performance of the power plant, were investigated (Ref
3). Carbon dioxide, carbon monoxide, water vapor, argon,
oxygen, helium, nitrogen, hydrogen and air are the gases
considered in the study (Ref 2). In this paper only the results
for carbon dioxide are reported.
with high efficiency, is steam injection, which is also
employed for emissions abatement.
CONSTANT PRESSURE SPECJRC HEAT FOR CO2
1400
1300
1200
1100
1000
900
800
700
600
200
400
600
800
1000
1200
1400
1600
1800
2000
TEMPERATURE (h2
Figure 2. Constant pressure specific beat of carbon dioxide
Pressures are in atmospheres
2
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REGIONS OF REAL AND IDEAL BEHAVIOUR FOR CARBON DIOXIDE
10 —
9
8
um.
REAL REGION
REAL REGION
7
6
5
co
tu
tc
ct.
4 —
3
IDEAL REGION
2
1
0
400
200
600
800
1000
1200
1400
1600
/73WPER4TURE pg
Figure 3. Ideal and real regions for carbon dioxide
THERMAL EFROSYCY POWER OF GAS MESHES.
45.00
40.00
S. 35.00
s•
Lit
(..) 30.00
•
.
as I a "I
•
.
—I 25.00
ct
•
• SIMPLE CYCLE
REGENERATIVE CYCLE
20.00
• STEAM INJECTED CYCLE
15.00
10.00
• 114TERCO0LED CYCLE
•
0.000
50.000
100.000
150.000
200.000
250.000
POWER OUTPUT (MW)
Figure 4. Thermal efficiency and power of gas turbine cycles
With the passage of time, turbine entry
temperatures have increased continuously (Figure 5). In
order to increase the life of the components and improve
engine reliability and availability, these temperatures are
kept 200 -300°C below the maximum temperatures employed
in advanced aeroengines. However they are well over the
maximum metal temperature, represented by the continuous
line in figure 5. Hence special attention is always paid to
blade cooling, one of the critical areas in a new gas turbine
development, and to the materials employed
In the most advanced gas turbines a mass flow as
high as 20% or the main flow is used for blade cooling. It
must also be pointed out that although heat-resistant
materials represent only 2% of the cost of an advanced
combined-cycle power plant, they play an extremely
important role by enabling high operating temperatures.
3
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TURBINE BURY TEMPERATURE OF INDUSTRIAL GAS TURBINES
1600
1500
1400
1— 1200
C
DO
RD
8
1100
10 0 0
PD
009
Ko g
0 0
0 C 0 0 0
0
8
00
°
BE
00
0
0
0
o
Watk
0
0
oo
0g
0
%rie
M252
D
90
U500
S-816
oe
8 01
0
0 0
El 6
0
1N-738
n
ot" 1
0
0 0 0 0
°PE
a Ogu n°
0
g 1300
UI
EgE1009g 1010
o 0
° °
o a
0 0 0 El
DS GTD-111
Sngle Oystal
woo
Metal temperature
900
1945
1950
1955
1960
1965
1970
1975
1980
1985
1990
1995
YEAR OF AVAILAINUTY
Figure 5. Turbine entry temperature and metal temperature of gas turbines.
The most common closed cycle is the regenerative
type, many engines incorporating intercooling. The
intercooler increases the specific power of an engine and it
makes the regenerator more useful, because the compressor
discharge temperature is reduced.
Closed cycle turbine entry temperatures are
approximately 1100 K, much lower than those employed in
open cycles. This is due to the use of a heat exchanger as the
source of heat addition. This low temperature coupled with
the use of the regenerator results in gas turbines with very
low design pressure ratio. The penalties associated with
these characteristics can be partly overcome by employing a
working fluid with properties better suited to that design.
This can be either a pure substance or a mixture of different
gases.
polytropic efficiency 87%, low pressure turbine polytropic
efficiency 87%, free power turbine polytropic efficiency 87%.
The regenerator effectiveness is 80%. The pressure losses of
the heat exchangers, ducts, etc. are: 0.5% inlet, 3% each side
of the regenerator, 5% the bottoming cycle heat exchanger,
and 2% the precooler. A customer bleed of 1% is considered,
as well as shaft friction losses of 0.5%, and a power turbine
loss of 2%. For combined cycles the discharge temperature
of the heat recovery boiler is 420 K, in line with current
practice.
With the parameters selected above the design
performance of a gas turbine fitted with an optimised simple
steam turbine cycle was calculated in two situations: with
pure oxygen as an oxidiser and with oxygen and argon
(95%/5%). As stated before the small amount of Argon is a
result of using the cheaper separation process because the
cryogenic air separation plant operates at a higher
temperature. The presence of Argon makes a very minor
difference to the performance of the engine. Therefore, the
only results shown are those with Argon in the working fluid.
The performance prediction can be subdivided into
two parts: cycle optimisation plus selection and off-design
behaviour. For this purpose, a programme working on the
basic principles of TURBOMATCH (Ref. 4), the Cranfield
gas turbine performance simulation software, was generated.
BASELINE CYCLE PARAMETERS
In the semi-closed cycle heat addition takes place
using a combustor, hence the temperature will be tailor to
those employed in conventional open cycles. The working
fluid is imposed by the characteristics inherent to the cycle,
and the fuel is a low heating value gas obtained from coal.
Initially a turbine inlet temperature of 1473 K was
selected. 'This temperature may be considered as a peak
temperature. For base load 1300/1350 K seems more
reasonable. This is low, for a new gas turbine, when
compared to those shown in Figure 5 but the temperatures
illustrated here apply to natural gas or clean distillate fuels.
In the case investigated here, allowances have to be made to
the use of a lower quality fuel in the interest of longer time
between overhauls.
Once the turbine entry temperature has been
selected, the rest of the parameters of the cycle are typical of
the state-of-the-art in gas turbines: low pressure compressor
polytropic efficiency 90%, high pressure compressor
polytropic efficiency 89%, combustor efficiency 99.9%,
combustor pressure losses 5%, high pressure turbine
DESIGN PERFORMANCE OF THE CYCLE
The efficiency of the four gas turbine cycles
considered here is shown in the combined gas and steam
cycle form. They are: the engine with a simple cycle gas
turbine (figure 6), intercooled (figure 7), regenerative (figure
8), and regenerative & intercooled cycle (figure 9). They all
employ advanced full coverage film cooling in the turbine
vanes and blades.
4
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••
COMBINED CYCLE EFFICHWCY
roHr.6
malE:AF,He
y• % 4.r4 61 .
';9)Ed 40,1May
v:vak
BE,41w/A,A, 7
12
alD 0.45-0.465
11
0.435-0.45
10
114F:glik474 g 44, „
P.27M1 9
7%.
1.
cfl,
,;r.1(1111:
771,11:1111111:t4
rbzwr t„:
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NY
—140
M
INEKTfA
01011112
‘NliefabiLl`
'1411111121111614',.=
in
co
• 0.405-0.42
n 0.394.405
NEC PRESSURE RATIO
0. 0.375-0.39
6
:
6
:296
9
0 0.36-0.375
5
„
••
4ve
•
r /,A7
ria
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4
.404reA
A'
4
o-
7
• 0.42-0.435
(13
(31
LPC PRESSURE RATIO
0 0.345-0.36
e, 3
• 0.33-0.345
2
• 0.315-0.33
0
0 0.3-0.315
Figure 6. Thermal efficiency of combined semi-closed cycle gas turbine.
COMBINED & INTERCOOLED CYCLE EFFICIENCY
12
EMENE:, :
'Are
rfa ranngi4
S7Art:'? ,
mae:i
10
• 0.42.0.435
9
• 0.405-0.42
8
• 0.36-0.405
7
"HallaireISIMEIERam
in
---
0
HpC PRESSURE RAMO
0 0.375-0.38
6
0.36-0.375
5
0 0.3450.36
4
• 0.33-0.345
ItilielENEMIDEINIZIESIMEIBEEMIN 3
C,
• 0.435-0.45
11
2
2 0.315-0.33
12 0.3m.315
LPC PRESSURE RATIO
Figure 7. Thermal efficiency of combined & intercooled semi-closed cycle gas turbine.
5
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COMBINED & REGENERATED CYCLE EFRoVENCY
A:4Ria
12
till 0.45-0.465
./A 74,1/4, /922t.
11
0.435-0.45
Zt//,crAg11111if
'A'
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4.
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.72'4'47
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• 0.405-0.42
8
'7311j111111111111111111mr..;r,
v
M 0.42-0.435
9
7
A,
0.39-0.405
WPC PRESSURE RATIO
0 0.375-0.39
6
4.111111111111111111MIIIK
63 0.36-0.375
5
:).1,1 91111111111111111111
A.V
r' ' ; 74V 7: /7;
7Y A. Alig
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0 0.345-0.36
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2
MI 0.315-0.33
?".".
414 47
1(Z.
CO sr ID CO 1`. CO 0) 0 9—
0 0.3-0.315
LPC PRESSURE RATIO
Figure 8. Thermal efficiency of combined & regenerated semi-closed cycle gas turbine.
COMBINED & INTERCOOLED & REGENERA TB) CYCLE EFRCENCY
12
WAY/74',"
rizr
,
11
47,
,7 V 47 7
?
Hr L4r 4 7
•'<e'4VV/r AZ 'Y''
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.
•
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N 01 Nr 11) CO
10
MI 0.42-0.435
9
• 0.405-0.42
8
MI 0.39-0.405
7 moo PRESSURE RATIO
CO 0
0
0.375-0.39
6
D 0.36-0.375
5
13 0.345-0.36
4
,:'•
0.435-0.45
Eel 3
2
• 0.33-0.345
El 0.315-0.33
0 0.3.0.315
LPC PRESSURE RATIO
Figure 9. Thermal efficiency of combined & antercooled & regenerated semi-closed cycle gas turbine
The reheat cycle has not been considered here
because of the uncertainties and the complexity a second
combustor would introduce to the design. Another feature is
that all turbornachinery is specifically designed for carbon
dioxide. Early in the analysis it became clear that the
efficiency penalty associated with using existing
turbomachinery was of the order of ten percent (Reference 3).
This VMS deemed to be unacceptable.
The reason for showing only the performance of the
combined cycles, and not just the gas turbine, is that this is
the most attractive option for electricity generation. If the
The best efficiencies of the semi-closed cycles fitted
with a steam battening cycle are of the order of 43-46 %.
Conventional combined cycles, shown in figure 10, currently
achieve thermal efficiencies well in excess of 50%. This
difference of the order of 10% is mainly due to the large
quantity of energy required by the oxygen separation process.
Another, smaller, penalty is that due to the higher
compressor inlet temperature applicable to closed and
semiclosed cycles. Low overall pressure ratio semiclosed
cycles are further penalised by the power needed to compress
the excess CO2 to the disposal pressure of at least 60 bar.
6
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simple cycle with LPC and HPC pressure ratios of 6 and 8
respectively. This engine is the subject of the off-design
analysis described in the next section.
The high pressure ratios required by the carbon
dioxide Cycle indicate a requirement for a two spool gas
generator.
interest was in cogeneration, a single cycle gas turbine will
be the choice. Due to the very high temperature found at the
exit of power turbine, the cycle presented here can be used
for industries where the power to heat ratio is low.
The design point analysis carried out indicated that
the best efficiencies are achieved with the simple cycle and
the regenerated cycle. The cycle selected is therefore the
COMBNED CYCLE THERMAL EFRCIENCY
55.00
•
U
• a
. is
50.00
,..E.-
45.00
•
to • db •
•
•
NS
... • t__ e ...II . .., . • - Es
„I
• ••
.•I •
• ••
dig
is •
•
it
Mg
it
•
•
•
4
uj 0.00
35.00 •
30.00
25.00 11
0.000
100.000 200.000 300.000 400.000 500.000 600.000 700.000 800.000
POWER (MW)
Figure 10. Thermal efficiency & power of conventional combined cycle gas turbine.
OFF-DESIGN PERFORMANCE OF THE CYCLE
second one crossed the LPC surge line at a high pressure
shaft speed of 90.75%.
This handling problem can be avoided by using
variable stators. This is preferred to bleed valves discharging
to the exhaust because the high design pressure ratios mean
that they will have to be used at high powers. If there is a
requirement to operate at part load the consequent efficiency
penalty cannot be accepted. The present design however will
probably need to incorporate bleed valves for starting
purposes.
The nature of the compressor design, subsonic or
transonic, will have a great influence in the surge margin at
part load. Initially both low pressure compressor and high
pressure compressor were subsonic, the initial
turbomachinery design being described in Reference I. Then
the LPC was changed to an advanced transonic compressor,
with a large pressure ratio per stage. The LPC surge margin,
in this case decreased much more.
Real maps for a CO2 compressor should be
obtained by experiment to evaluate, more accurately, the real
behavior of the machine. In the simulations presented here
the maps employed were the same for air and for carbon
dioxide, naturally the correct non dimensional numbers were
used
The handling issues and the less attractive offdesign efficiency require further study. These features do not
seem to present any insuperable problems.
The off-design performance of a semi-closed cycle
combines the features of the off-design performance of an
open cycle and a closed cycle. Theoretically the compressor
inlet pressure can be varied. However, because water must
be removed as a liquid before the compressor inlet, changes
in this pressure are vay limited. The compressor inlet
temperature is limited by the precooler water temperature.
Hence the main control parameter is the turbine temperature.
Nevertheless the limited control in compressor inlet pressure
will allow small changes in power output without significant
changes in efficiency.
In table I, the performance of a two spool gas
generator with an independent power turbine is presented.
These results show that the efficiency reduction at part load
is larger than what would be expected from a conventional
open cycle gas turbine.
It is important to note that the high pressure ratios
required for a good CO2 cycle, will cause off-design
problems, specially in compressor handling. Splitting up the
gas generator into two spools will alleviate but not remove
these problems. •
The handling of an open cycle gas turbine was
compared to that of a semi-closed cycle one, both using a
high heating value fuel and with the same compressor
pressure ratios. The low pressure compressor (LPC) surge
margin of the first one fell from 25% to 9% when the high
pressure shall speed fell from 100% to 90%, while the
7
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TABLE I: Semi-Closed Cycle Performance
(Simple & ComMoe& Cycle)
Mass Flow (%)
LPC P.Rasio
IIPC P.Ratio
CONCLUSION
This paper describes the preliminary performance
analysis of a semi closed cycle gas turbine. The working
fluid is carbon dioxide and the fuel is low heating value gas
synthesised from coal. The selected configuration is a two
spool simple cycle gas generator with an independent power
turbine.
The predicted behaviour of this engine indicates
that an efficiency penalty has to be paid if improved
environment-friendliness becomes an important criterion
when plant is selected. Some off-design features are not so
attractive as those of conventional gas turbines and some
handling problems have been identified. These however do
not appear to present unstumountable barriers.
Rotational speeds and size also do not present any
extraordinary features. It is therefore expected that the
mechanical design of this machine will offer no more
problems that of a conventional one.
The main issue will be an economic one. It is very
difficult at this stage to make a realistic estimation of the cost
of a power plant of this type. The uncertainties in the
development budget, the number of machines to be built, etc.,
contribute to this difficulty. It is clear that such a machine
will be much more expensive than a conventional one. In the
balance will be its higher cost and lower efficiency against its
ability to use coal and its attractive environment features.
The prize offered by this type of engine is clean
electricity. There are many difficulties envisaged in the
development and construction of such a machine.
Furthermore its efficiency will be lower than a conventional
open cycle combined cycle gas turbine. Its development
depends on the willingness to pay a premium for clean
electricity.
100.00 93.01 84.13 74.74 65.94 58.40 51.47
6.00
5.70
5.27
4.80
4.35
3.96
3.57
8.00
7.70
7.39
7.07
6.75
6.45
6.17
LPC RPM (%)
100.0
96.8
93.3
90.0
87.2
84.7
81.8
FIPC RPM (%)
100.0
98.1
96.2
94.0
92.0
90.2
88.4
TET (K)
1473
1423
1373
1323
1273
1223
1173
Fuel (Kg Is)
10.48
9.18
7.81
6.50
5.36
4.41
3.62
ETA_SC (%)
26.22 25.38 23.84 21.73 19.12 16.43 13.43
ETA CC (%)
47.26 45.82 43.49 40.75 37.63 34.22 30.41
DEVELOPMENT ISSUES
A major factor that will affect the development of
the engine is the lack of a suitable knowledge database. Air
is the working fluid of almost all existing gas turbines.
Knowledge of operation with different working fluids is very
limited.
Several closed cycles have been in operation, some
with air, others with helium, mixtures of helium and xenon,
etc. However the total experience accumulated after some 50
years in base load power generation is around 1 million
hours. This appears to be a large number, but it is very
small compared with operational experience of some existing
gas turbine models, many having operated for tens of millions
of hours.
There will be three major areas of risk in the gas
turbine: the aerodynamic design, the combustor design and
blade cooling. The development of an extensive theoretical
and experimental knowledge database will be necessary to
support the development of the above. Improvement and
validation of Computational Fluid Dynamic codes will be
necessary to handle suitable gases and their mixtures. Some
basic design criteria will also need updating, due to the
different gas properties.
Expensive compressor caarade rigs, and at a later
stage compressor test stands will be required to examine
aerofoil performance in exotic fluids. In particular in the hot
section, it is probable that new materials and coatings will be
required to cope with the idiosyncracies of the working fluid
and the fuel employed.
A similar lack of experience will make the
development of the combustor a high risk item. There is no
expertise in burning a low heating value gas with pure
oxygen in a CO2 environment With the high temperatures
expected, and a requirement for combustion efficiency of
100%, extensive tests must be carried out to ensure proper
combustion, as well as the absence of pollutants.
In the case of cooling technology the current state of
the art is also based on the large number of experiments
carried out, in other words, on existing experience. In
existing closed cycles, where different gases were employed,
no cooling was used, because turbine inlet temperatures were
low. If a competitive machine is to be designed,
temperatures will need to be be high to achieve good
efficiencies. Therefore blade cooling can not be avoided.
ACKNOWLEDGMENT
The authors acknowledge the financial support
obtained from the C.E.C. in carrying out this task They are
also grateful to their colleagues in LT.P. and Cranfield for
their continuous support and encouragement.
REFERENCES
1
Hunter, I.H.
Design of Turbomachinery for Closed
and Semiclosed Gas Turbine Cycles.
M.Sc. Thesis, Cranfield University, 1994
2
Keenan, J.H., Kaye, J. and Chao, J.
Gas Tables, International Version,
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