Note: Descriptions are shown in the official language in which they were submitted.
... 2
BACKGROUND AND SUMMARY OF THE INVENTION
This invention relates to a closed compression gas-turbine and superheated/reheated steam-
or vapor-turbine combined-cycle power-producing system. In such a configuration, according
to this invention, it is ultimately important that the exhaust gas from a said gas turbine on any
terms has very, but optimally, high temperature, higher than 873 K (600~C or 1112~F), so that
cycle thermal efficiency upwards of 60% can be achieved.
It is generally known that the highest possible efficiency of converting thermal energy (obtain-
ed in the combustion process of fuel) into mechanical or electrical energy, that is, the thermal
efficiency, can be obtained in a combined gas- and steam-turbines cycle, where exhaust gas
from the gas-turbine cycle serves as a heat source for a bottoming steam-turbine cycle. The
heat from the exhaust gas is recovered in a heat recovery boiler, or heat recovery steam
generator (HRSG), which generally consists of: a condensate/feedwater economizer, an
evaporator, a steam superheater and a steam reheater. In the conventional practice, the
temperature of superheated/reheated steam (which is not necessarily the same) is set at the
maximum permitted by the materials employed in the boiler superheaters, reheaters, steam
piping, valves and steam turbines, which is in the range of 1 000~F (538~C) to 1 050~F (566~C).
The boiler can evaporate water, ammonia, Freon or another liquid or a mixture, to form super-
heated vapor for expansion through a vapor turbine, such as a steam turbine. In practice,
however, only water has so far been proved as the most desirable working fluid, regarding its
thermodynamic properties, cost of production and chemical composition stability at said high
temperatures.
The other fact that is also known, but which is usually not enough emphasized, is that a rehe-
at-gas-turbine combined cycle enables obtaining higher combined-cycle efficiency than other-
~
...3
wise obtainable from simple-gas-turbine combined cycles. This happens due to higher possi-
ble exhaust-gas temperatures in case of the gas-reheat, depending on the chosen compress-
or pressure ratio. Higher exhaust-gas temperature means that greater power output can be
obtained from the steam-turbine part of the combined cycle, because of either larger achieva-
ble mass flow rate of water-steam, or higher achievable superheated/reheated steam maxi-
mum temperature, or because of the combined effect of both. For combined cycles the gas-
-turbine-exhaust-gas temperature can be controlled by the selection of second-turbine inlet
temperature and expansion (compression) ratio. This in turns allows control over the efficien-
cy of the bottoming steam-cycle. This is a very important fact, of which has been taken main
advantage in this invention.
It is also generally known that the combustion air required in the open gas-turbine cycle is
provided by compressors. Part of the compressor air is frequently branched-off in order to
cor,l, ibute to the cooling of the gas-turbine blading, particularly nowadays when extremely
high gas-turbine-inlet temperature (up to ~2500-2600~F) is an indispensable requirement for
achieving a high gas-turbine-cycle thermal efficiency. Cooling of the gas-turbine vanes and
rotors allows the gas-turbine inlet temperature to be increased beyond the temperature at
which the gas-turbine material can be used without cooling, thus increasing the cycle thermal
efficiency and power output. However, the cooling air itself is a detriment to cycle thermal effi-
ciency because it reduces hot-gas-path temperature, increases the turbine-airfoils's irreversi-
ble pressure losses and introduces irreversible pressure losses due to the nonideal mixing of
the two streams with very different velocity vectors. Furthermore, the cooling air must be com-
pressed to pressures significantly higher than that of the hot-gas-path pressure at the location
it is injected, in order to assure a sufficient cooling-air flow rate during all operating conditions.
Some of this pressure is recovered by the gas turbine.
Since the cooling air is usually compressed to very high pressures and temperatures, it is too
~J~ ...4
much hot to be efficiently used directly in the process of the gas-turbine rotor cooling. So, it is
general practice today to recover the rotor-cooling-air thermal energy by the bottoming steam-
-cycle, which produces additional low-pressure steam with the heat, thus slightly improving
combined-cycle thermal efficiency. There is another proposal of recovering the rotor-cooling-
-air heat by exchanging it with the incoming natural gas fuel (since natural gas is usually used
as a fuel in open gas-turbine cycles). This returns the rotor-cooling-air heat back to the com-
bustion chamber, which than requires less fuel to achieve the desired gas-turbine-rotor inlet
temperature. Undoubtedly the best proposal is the use of a closed-loop steam cooling, ins-
tead of use of the cooling air, to maintain gas-turbine-material temperatures at an acceptable
level. In combined cycles this steam would be taken from the bottoming steam-cycle, precisely
from the exit of the high-pressure steam turbine. This approach to gas-turbine cooling relies
solely on convective heat transfer. The associated reduction in hot-gas-path temperature is
minimized, since the convective heat flux across the airfoils is relatively small. Typically, mix-
ing of the first-vane cooling air with the hot-gas stream reduces the hot-gas-path temperature
approximately 100~F to 150~F (56~C to 83~C). For closed-loop steam cooling however, this
temperature reduction is only about 10~F to 1 5~F (6~C to 8~C), or one tenth of the temperature
reduction of conventional air-cooling techniques.
All combined-cycle configurations mentioned so far, are open-type cycles with natural gas as
a fuel and thermal efficiencies based on the International Standards Organization (ISO) refe-
rence conditions (60% humidity, 14.7-psi/59~F ambient air) and lower heating value (LHV) of
natural gas. This means that water vapor, which was formed in the combustion process, is not
condensed to liquid and, consequently, the latent heat of condensation is carried away by
vapor in the exhaust gases.
According to information from Volume 117 of "Journal of Engineering for Gas Turbines and
Power" from October 1995, the United States Department of Energy is continuously working
-
~ ~4~
... 5
on the 8-year Advanced Turbine Systems Program, in cooperation with some worldwide lea-
ding co,npanies in the gas-turbine design and manufacturing (Westinghouse). The target of
the rroy,dm is to develop an innovative natural gas-fired advanced-turbine open-combined
cycle, which, in combination with increased firing temperature, use of advanced materials,
increased component efficiencies and reduced cooling-air usage, has the potential of achie-
ving a lower-heating-value combined-cycle efficiency of 60% by the year 2000. However, ac-
cording to this invention this goal is achievable almost today, without increasing component
efficiencies or reducing cooling-air usage, provided that minimum gas-turbine-inlet temperatu-
re of 1625 K (1352~C or 2465.6 ~F) can be used, in a closed-combined-cycle configuration.
This invention proposes a closed-combined gas- and steam/vapor- turbine cycle with high-
-temperature exhaust gas from the gas turbine, without gas-reheats, or with one, two or more
gas-reheats. It is an object of this invention to point out that, depending on a chosen closed-
-cycle configuration, parameters of live/reheated steam (or vapor), number of steam-reheats
and condensation temperature, it exists an optimal exhaust-gas temperature, coupled with an
appropriate compressor-inlet temperature, as well as an optimal compression pressure ratio,
for achieving maximum closed-combined-cycle thermal efficiency. It has been found that the
best closed-combined-cycle configuration is a single-reheat gas-turbine cycle with a single-re-
heat bottoming steam-turbine cycle. In accordance with general engineering practice, it is
assumed that all turbine-cooling gas is cooled after its compression (prior to its injection into
the gas-turbine vanes and blades) with saturated water from the steam-cycle, thus producing
an additional amount of intermediate-pressure saturated steam. Assuming use of the closed-
-loop steam cooling of the gas turbine(s), maximum thermal efficiency of 62.5 to 63.5 % can
be achieved with a single-reheat-gas-turbine closed-combined cycle, or even 64 to 65% with a
dual-reheat-gas-turbine closed-combined cycle, at gas-turbine-inlet temperature of 1625 K
(1352~C). In all considered closed-combined-cycle configurations a bottoming steam-cycle is
assumed, with conventional maximum superheated/reheated-steam temperature of 540~C
'~ A~
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...6
(1004~F) and maximum superheated-steam pressure of 175 bars (2538 psi).
According to this invention, closed-combined-cycle concept can be ideally applied to repower
existing steam-turbine power-plants. Repowering is a term that means replacement of a ste-
am boiler with a heat recovery steam generator, which recovers heat from exhaust gases of a
gas-turbine open cycle. It is also possible and desirable to perform repowering of the existing
steam power-plants within a closed-combined-cycle configuration, as proposed according to
this invention, with an additional important remark: here there is no longer need for regenera-
tive condensate/feedwater heaters, which have been conventionally used for improving steam
-cycle thermal efficiency. In a closed-combined cycle, as it was said earlier, it is important that
the compressor-inlet temperature has its lowest possible (or optimally low) value, which is
usually bellow 100~C (212~F). While use of regenerative heaters in the bottoming steam-cycle
of a closed-combined cycle improves thermal effficiency of the steam-cycle alone, it does not
increase the total combined-cycle efficiency. First reason is that gas-turbine-cycle efficiency
decreases in the same time, because of greater work of compression required due to increa-
sed compressor-inlet temperature (although heat input needed is smaller, because compres-
sor-outlet temperature is also higher). Second reason is that the mass flow rate of the steam
produced does not increase, because less energy can be recovered from the exhaust gas,
since higher feedwater temperature causes the higher compressor-inlet temperature (although
in the same time less energy is needed for the same steam production). In other words, rege-
nerative heaters more decrease gas-turbine-cycle efficiency than they increase steam-turbine-
cycle effficiency in a closed-combined cycle. Consequently, it is not recommendable to use
them in this type of combined-cycle from the viewpoint of both thermodynamic efficiency and
investment cost, according to this invention.
A further object of this invention is to emphasize the fact that any type of fuel can be used in
the closed-combined-cycle gas-furnace. This is particularly important with respect to coal, as
~,
~ ~ B4 ~ ...7
a most abundant fossil fuel in the world. No coal gasification is needed, which has been pro-
posed for indirect usage of coal in open-combined-cycle configurations. Instead, the conventi-
onal method of pulverized coal combustion can be used in the furnace that mostly resembles
a conventional steam boiler, regarding its structure and function. Coal, or any other solid fuel,
such as solid waste materials, can be burned even in a layer on the conveyer grid. The only
difference between this furnace and a conventional steam boiler is that the water-steam mixtu-
re is replaced with a gaseous working fluid, such as dry air, with one gas heater and with (or
without) at least one gas reheater. Because of the much higher maximum allowable and achi-
evable gas temperatures in comparison to restricted maximum superheated/reheated steam
temperatures, the mean heat-exchange temperature of the furnace is also much higher, which
results in a higher heat-exchange effectiveness of the furnace in comparison to that of a con-
ventional steam boiler. Pressurizing of the working gas can lead to rather high heat-transfer
coeffficients and, consequently, the tube-wall temperature is not much higher than the maxi-
mum gas-turbine temperature, the differences amounting to only 30 to 60~C. Experience with
existing closed-cycle gas-turbine power-plants shows that uniform turbine-inlet-temperature
distribution in the different small gas-heater tubes can be achieved without difficulty. Since the
tubes are subjected to high wall temperatures, good combustion at all loads is ensured. Pulve-
rized-coal-fired sets can work at a wide load range, down to 20% of the normal load, without
additional oil- or gas-flames, which are only necessary for starting. Also, it is possible to pre-
heat ambient air, necessary for complete combustion of coal, to considerably higher tempera-
tures, which also contributes to higher heat-exchange effectiveness of the gas-furnace in com-
parison to that of a conventional steam boiler.
Another object of this invention is to remind of the fact that not only dry air should be used as
a working fluid in a closed combined cycle. For instance, helium can be used, as in the case
of a gas turbine with a helium-cooled nuclear reactor. Special physical properties of helium
favor its use in the closed-circuit systems. Higher a specific-heats ratio of helium (1.667) in
C
~4 ~
...8
comparison to that of dry air (1.4), influences that much lower optimum compression pressure
ratio (CPR) is necessary for obtaining maximum closed-combined-cycle thermal efficiency,
than in case of dry air as a working medium. In addition, helium has many-times-greater
thermal conductivity (about 6 times) and constant-pressure specific heat (about 5 times) than
those of dry air. This means that, with the same mass flow rate of gas, it is possible to obtain
about 5 times greater power output with helium as a working fluid than with dry air, at approxi-
mately the same or even better closed-combined-cycle thermal efficiency. Also, this means
that heat-transfer coefficients are higher (about twice), which leads to small heat-exchange
surfaces, especially at appropriate elevated closed-cycle pressures of 40 to 70 at (550 to
1000 psi). As helium has much lower viscosity than air, pressure losses are smaller compared
with air cycles. Very high temperatures cannot be dealt with in steam- or air-cycles, due to the
fact that most high-temperature and refractory materials have poor oxidation properties. A
closed-cycle working with helium (or any other monatomic inert gas, such as neon or argon)
offers a unique possibility of using such high-temperature materials as, for instance, molybde-
num and its alloys and niobium, since no oxidation danger exists. In neutral atmosphere these
materials can support high stresses at very elevated temperatures.
Because of its favorable thermophysical properties, helium can be an ideal working gas for
repowering of the existing and future steam-turbine power-plants. With an appropriate mass
flow rate of helium, literally any steam-turbine plant can be transformed into a highly-efficient
closed-combined cycle. Also, helium closed-combined cycle can be used in high-temperature
gas~ooled reactors instead of usually used intercooled-recuperated closed-gas-turbine cycle,
because the proposed configuration has much higher thermal efficiency and power output.
Another great possibility, with respect to repowering, is the use of hydrogen. Although it is
very flammable, hydrogen is not corrosive and, in addition, has about 3 times greater constant
-pressure specific heat than helium, that is, 3 times greater heat capacity. This means that it is
'' ~ '
2' J~ ... 9
possible to achieve the same power-plant output with 3 times smaller mass flow-rate of hydro-
gen as a working fluid at nearly the same cycle thermal efficiency, than if helium is used as a
working fluid. Similarly as in case of helium, with hydrogen gas-turbine-inlet temperature can
be raised to the ultimate limit permitted by stability of used high-temperature material.
Nitrogen has very similar thermophysical properties to those of dry air, but it is chemically
stable and nonaggressive/noncorrosive and, therefore, more desirable as a working fluid.
Carbon-dioxide (a specific-heats ratio of 1.29) is another possibility as a working gas, but its
closed cycle typically needs impractically high CPR for maximum thermal efficiency. Also, it is
increasingly chemically aggressive/corrosive at high temperatures. Dissociation (decompositi-
on) of carbon-dioxide to carbon-monoxide and oxygen at high temperatures will be at negligib-
le level at usually used gas-turbine-inlet temperatures. Carbon-dioxide closed cycle shows a
closer approach to the steam cycle in many respects, due to distortion at higher pressures. At
higher pressures the distortion of the pressure-volume behavior reduces the compressor work
needed and thus boosts the useful specific power output. However, in the same time, conside-
rable increase of carbon-dioxide specific heat at higher pressures results in increased fuel
consumption for the same turbine-inlet temperature. Higher efficiency can be achieved using
helium-carbon-dioxide mixture as a working fluid.
Another way of decreasing maximum absolute gas pressure ahead of the closed-cycle gas
turbine, other than use of a monatomic gas as a working fluid, is use of the lower-pressure-
-level gas compressor and turbine (with the same required CPR), so that a certain vacuum
exists in the heat recovery steam generator (HRSG), from the gas-turbine exit to the compre-
ssor inlet. This interesting solution would be complicated by additional problem of eventual
leakage of ambient air into the closed gas circuit through the HRSG; so in this case major
concern would be sealing of the parts of installation exposed to such a vacuum.
P, ~
~ ,,
... 10
Since the proposed combined-cycle operates in a closed circuit, it is possible to apply cooling
of the gas turbine with water or steam circulating in a closed loop, with mixing of hot gas and
thus superheated steam at the exit of gas-turbine vanes and blades. This solution represents
an intermediate solution between usually used gas-cooling system and the best option of clo-
sed-loop steam cooling without mixing of steam and hot gas. While the latter solution uses so-
lely convective heat-transfer to cool gas-turbine vanes (and blades), this intermediate solution
removes the heat by evaporative coolingj which is much more effective way of cooling than
convection. After mixing with the hot gas, thus obtained steam performs work expanding thro-
ugh the gas turbine and then goes to the HRSG in which it is cooled and partially condensed
by the water from the bottoming steam-cycle. The temperature of eventual condensation de-
pends on the partial pressure of water-vapor in the gas-vapor mixture, that is, the amount of
vapor in the mixture, and the total pressure of the mixture in the HRSG. Complete condensati-
on of the cooling water is accomplished in an additional condenser, located between the exit
of the heat exchanger (HRSG) and the inlet of the compressor, which serves to separate
vapor from the gas. Condensed water (or steam produced in the HRSG) is then again pumped
to the gas turbine(s). Because of corrosion danger due to condensation, if air or carbon-dioxi-
de is used as a working medium, a corrosive-resistant material should be used for manufactu-
ring of the additional condenser, as well as for a part of the HRSG. With use of helium or
hydrogen, probability of corrosion due to cooling-water condensation is almost negligible.
Water or steam can be added in even larger quantities in order to increase gas-turbine power
output, as in the case of open cycles. With the appropriate choice of enough high last-gas-
turbine outlet pressure, the heat of condensation of this additional water can be used by the
bottoming steam-cycle. In such a case, however, the gas-turbine circuit would be exposed to
very high pressure between the compressor exit and the gas-turbine inlet.
Finally, still another object of this invention is to show that solution of the combined cycle with
high-temperature exhaust gas is not limited only to a closed-, but it can also be applied to an
~
... 11
open-combined cycle. Here, high temperature of exhaust gas is also important because it
enables higher cycle power output. Corresponding low temperature of exhaust gas at the exit
of the HRSG is important because it provides necessary condition for eventual condensation
of water-vapor from the flue gas, which was formed or added in the combustion process of
natural gas, a usual fuel in open gas-turbine cycles. This means that, because it exists a
possibility of recovering a certain amount of the latent heat of condensation, thermal efficiency
of such an open-combined cycle could be even higher. However, in the same time if the
HRSG-outlet temperature is reduced to the water-condensation point, eventual corrosion may
result, either in the heat exchanger (HRSG) or in the exhaust stack. This can be even worse if
there is any sulfur in the fuel, because sulfur oxides, in combination with excess air, raise the
dewpoint temperature very significantly, producing liquid sulfuric acid at well above the dew-
point temperature. Exhaust stacks and heat exchangers can be quickly corroded in this way.
So, according to this invention, the solution with high-temperature gas-turbine exhaust gas
can be applied to open-combined cycles because of the gain in cycle thermal efficiency, but in
the same time an expensive corrosion-resistant material should be used for manufacturing of
the heat exchanger and the exhaust stack in this case.
~
... 12
BRIEF DESCRIPTION OF THE DRAWINGS
In the first five drawings (FIGS. 1, 2, 3, 4 and 5), an illustrative embodiment of a preferred
embodiment of the subject of the invention is shown diagrammatically. Parts that are not
essential to the invention, such as, for example, electrical drives of the condensate/feedwater
pumps and the forced draft fan, are not shown. The direction of flow of the various working
media is marked with arrows: wide solid line denotes the gaseous working fluid flow, narrow
solid line denotes the steam-water flow, dashed line denotes the combustion air flow and triple
solid line denotes the flue-gases flow.
FIG. 1 is a circuit diagram of a compression gas turbine and superheated/reheated steam
closed-combined cycle without gas-reheats, with high-temperature exhaust gas being conduc-
ted to a heat recovery means.
Consequently, FIGS. 2 and 3 are circuit diagrams of the same said closed-combined cycle
with one and two reheats of the working gas, respectively.
FIGS. 4 and 5 are circuit diagrams of the same closed-combined cycle without gas-reheats
and with single gas-reheat, respectively, with gas-turbine cooling performed by injected water,
circulating in a closed circuit.
FIG. 6 is a temperature-versus-heat-exchanger-length graph in a heat recovery steam gene-
rator (HRSG), for three different exhaust-gas temperatures at the inlet and a steam cycle with
superheated and reheated steam.
FIGS. 7 and 8 are thermal-efficiency-versus-specific-power graphs of a closed-combined cyc-
... 13
le without gas-reheats, with air and helium as a working medium, respectively, for three diffe-
rent gas-turbine-inlet temperatures.
FIGS. 9 and 10 are thermal-efficiency-versus-specific-power graphs of a closed-combined
cycle with single gas-reheat, with air and helium as a working medium, respectively, for the
same range of gas-turbine-inlet temperatures.
FIGS. 11 and 12 are thermal-efficiency-versus-specific-power graphs of a closed-combined
cycle with two gas-reheats, with air and helium as a working medium, respectively, for the
same range of gas-turbine-inlet temperatures.
FIGS. 13 and 14 are thermal-efficiency-versus-specific-power graphs of a hydrogen closed-
-combined cycle without gas-reheats and with single gas-reheat, respectively, for the same
range of gas-turbine-inlet temperatures.
FIGS. 15 and 16 are thermal-efficiency-versus-pressure-ratio graphs of a closed-combined
cycle with single gas-reheat and closed-loop steam cooling of the gas turbine, using helium
and hydrogen as a working medium, respectively, for the same range of gas-turbine-inlet
temperatures.
Consequently, FIGS. 17 and 18 are thermal-efficiency-versus-pressure-ratio graphs of a clo-
sed-combined cycle with two gas-reheats and closed-loop steam cooling of the gas turbine,
using helium and hydrogen as a working medium, respectively, for the same range of gas-tur-
bine-inlet temperatures.
... 14
DESCRIPTION OF THE PREFERRED EMBODIMENTS
In FIG. 1 a circuit diagram of the combined-closed cycle without gas-reheats is shown, inclu-
ding: a gas-furnace group 1, a gas-turbine group 2, a heat-recovery-means group 3 and a
steam-turbine group 4.
The gas-furnace group 1 consists of: a solid fuel inlet line 5 (preferably coal, pulverized or in
layer, or any solid or liquid waste fuel), a forced draft fan 6 (or induced draft fan, or both of
them if needed) with an air-preheater 7 that preheats ambient air needed for combustion, a
working gas heater 8 and a flue-gases chimney 11. As it was mentioned before, a closed-cyc-
le gas furnace mostly resembles a conventional steam boiler, structurally and functionally,
with higher heat-exchange effectiveness due to higher mean heat-exchange temperature and
higher possible preheating temperature of combustion air, in comparison to those of a conven-
tional steam boiler.
The gas-turbine group 2 includes: a gas compressor 12 and a gas turbine 13, with the said
gas turbine driving both a load 16, preferably an electric generator, and said gas compressor
12 directly by a shaft.
The heat-recovery-means group 3 includes a heat recovery steam generator (HRSG) for the
conventional superheated/reheated high-pressure steam generation, comprising: a conden-
sate/feedwater economizer 17, an evaporator 19, a steam drum 18 for separation of water and
vapor, a steam superheater 20 and a steam reheater 21. For achieving of maximum cycle
efficiency, it is most important to expose the high-temperature section of the superheater 20,
or that of the reheater 21, or both, to an optimum exhaust-gas temperature, which may be very
high (in the vicinity of 1500~F). Therefore, an improved high-alloyed steel may be required as
... 15
a construction material for the said high-temperature sections of the superheater 20 and/or
reheater 21. The economizer 17 may be needed larger than that of a conventional HRSG, in
order to enable achieving of very small needed temperature differences between the conden-
sate/feedwater at the inlet of the economizer 17 and the gas at the outlet of the HRSG 3, that
is, at the inlet of the compressor 12. In addition, the heat-recovery-means group includes: an
additional small heat exchanger 30 (that can also be an integral part of the HRSG), which re-
covers heat from the gas-turbine-cooling gas after its discharge from the compressor 12, an
additional intermediate-pressure steam drum 32 (it can also be used an existing intermediate-
-pressure steam drum in the HRSG, which is not shown in the Figures) for separation of pha-
ses and an additional small-size steam reheater 31, which is used to collect and superheat
the additional intermediate-pressure saturated steam (for this purpose enough larger existing
steam reheater 21 can also be used instead). After cooling in the heat exchanger 30, the gas-
-turbine-cooling gas is being pumped to vanes and blades of the gas turbine 13 in order to
cool them, after which operation it forms a mixture with the hot-gas stream and expands thro-
ugh the gas turbine 13, before it enters the HRSG.
The steam-turbine group 4 includes: a high-pressure steam turbine 22, an intermediate-
-pressure steam turbine 23, a low-pressure steam turbine 24, a condenser means 25 that
condenses expanded steam by cooling water, a condensate pump 27, a feedwater pump 28
and a feedwater tank with deaerator 26. It has to be emphasized that function of the feedwater
tank with deaerator 26 is not, in this case, to preheat the condensate/feedwater, but only to
remove oxygen from it, prior to its further chemical treatment, as well as to store it. Therefore,
only a very small amount of steam extracted from the outlet of the intermediate-pressure turbi-
ne 23 (as depicted in FIG. 1) should be used to drive off all dissolved gases from the conden-
sate/feedwater. If it is economically justified by significantly improved cycle thermal efficiency,
the removal of dissolved gases could be done only by chemical means, without extracting any
amount of steam from the turbine, thus keeping the condensate/feedwater temperature at as-
. ~
... 1 6
-low-as-possible level. For the same reason, regenerative condensate/ feedwater heaters are
not necess~ry in the steam-turbine group 4 of a closed-combined cycle, as it was emphasized
earlier in the text.
For the sake of simplicity and space savings, in FIG. 1 (as well as in FIGS. 2, 3, 4 and 5) it is
depicted that the same load 16 (an electric generator) is driven by the gas turbine 13 at one
end and by the steam turbines 22, 23 and 24 at the other, a so-called single-shaft power-train
concept. In practice, the traditional dual-shaft power-train concept can also be used, with se-
parated loads (electric generators) driven by the gas turbine and steam turbines, respectively.
The configuration shown by the circuit diagram in FIG. 2 relates to a combined-closed cycle
with single gas-reheat. The circuit diagram shown corresponds substantially to the circuit
diagram shown in FIG. 1, so that the corresponding reference numbers have been used in
each case for the features that coincide and for which, therefore, the explanations given with
respect to FIG. 1 apply. The essential difference relative to the previously explained configu-
ration in FIG. 1 is the addition of one more gas turbine 14 in the gas-turbine group 2, which
expands the gas additionally heated, that is, reheated, in a gas reheater 9, which is supple-
mentally incorporated in the gas-furnace group 1. As it was said earlier, the introduction of the
gas-reheat additionally improves combined-cycle thermal efficiency.
In FIG. 2 the shaft 29, which connects the two gas turbines 13 and 14, enables the entire
expansion pressure ratio to be adequately divided between the two said gas turbines, so that
an optimal gas turbine exhaust-gas temperature at the inlet of the HRSG is achieved. In
practice, the shaft 29 could be omitted and thus the gas turbine 13 would be the driving
turbine of the compressor 12, whereas the gas turbine 14 would be the free power turbine,
which drives the load 16. This arrangement is better from the standpoint of the load changes,
for it enables that the compressor can be driven at constant rotational speed, regardless from
... 17
eventual changes in the load demand, but, in the same time, it may lead to slightly lower cycle
thermal effficiency, because an optimal gas-turbine exhaust-gas temperature may not be achi-
evable in this case.
The configuration shown by the circuit diagram in FIG. 3 relates to a combined-closed cycle
with two gas-reheats. The circuit diagram shown substantially coincides with the diagrams
shown in and explained in FIGS. 1 and 2, respectively, so that the corresponding reference
numbers have been used in each case for the coinciding features, for which the explanations
given with respect to FIGS. 1 and 2, respectively, apply. The essential difference, relative to
the previously explained configuration in FIG. 2, is the addition of another gas turbine 15 in
the gas-turbine group 2, which expands gas additionally reheated in the second gas reheater
10, which is supplementally incorporated in the gas-furnace group 1. According to this inventi-
on, the maximum economical number of gas-reheats in a closed-combined-cycle configuration
is one, considering the maximum value of cycle thermal efficiency and the rate of the specific-
-power increase. If closed-loop convective steam-cooling system is used, than the maximum
economical number of gas-reheats is two, taking into account the maximum value of cycle
thermal efficiency and the rates of the cycle-efficiency and the specific-power increases,
according to this invention.
The configuration shown by the circuit diagram in FIG. 4 relates to a combined-closed cycle
without gas-reheats and with water-cooling of the gas-turbine's airfoils. The circuit diagram
shown corresponds substantially to the circuit diagram shown in FIG. 1, so that the correspon-
ding reference numbers have been used in each case for the features that coincide and for
which, therefore, the explanations given with respect to FIG. 1 apply. The essential difference
relative to the previously explained configuration in FIG. 1 is the substitution of conventional
gas-cooling system from the heat-recovery-means group 3 with water-cooling system, which
includes: a separated part 33 of the HRSG for eventual partial condensation of cooling steam
... 18
obtained by evaporation of the cooling water through the gas-turbine's airfoils, an additional
condenser 34 for complete condensation of the cooling steam, as well as for additional cool-
ing of the gas prior to its entrance in the compressor 12, and a cooling-water pump 35 for
pumping the condensed cooling water back into the gas turbine 13. After mixing with the hot-
-gas stream, the cooling steam expands in the said gas turbine producing additional work and
then enters the HRSG in which the heat removed from the gas-turbine is being recovered by
the bottoming steam-cycle.
The configuration shown by the circuit diagram in FIG. 5 relates to a combined-closed cycle
with single-gas-reheat and with water-cooling of the gas-turbine's airfoils. The circuit diagram
shown corresponds substantially to the circuit diagrams shown in FIGS. 2 and 4, so that the
corresponding reference numbers have been used in each case for the features that coincide
and for which, therefore, the explanations given with respect to FIGS. 2 and 4 apply. The only
difference relative to the previously explained configurations in FIGS. 2 and 4 is the addition
of a pressure-control valve 36 in the water-cooling system from the heat-recovery-means gro-
up 3, which serves to eventually adjust much higher pressure of the injected cooling water
with lower inlet pressure of the second gas turbine 14. With this configuration, even greater
work of cooling-steam expansion is obtained, but an additional heat is needed to reheat the
gas-steam mixture to designed gas-turbine-inlet temperature before its entrance in the second
gas-turbine 14. Instead of cooling water, it is possible to use cooling-steam from the bottoming
steam-cycle, raised in the heat recovery steam generator 3 (intermediate-pressure saturated
steam or cold-reheat steam). Also, it is possible to use larger amounts of steam (or water)
added in the gas turbine(s) for the purposes of both gas-turbines cooling and power-output
increase. All this additional steam can be completely condensed by the bottoming steam-cycle
(even without the additional condenser 34) if an appropriate pressure is selected at the exit of
the last gas turbine. However, in this case an extremely high pressure at the first-gas-turbine
inlet may cause unacceptable high stresses of the gas-turbine material.
-;G
... 19
THERMODYNAMIC ANALYSIS
The thermodynamic analysis of the heat-rate advantage of this invention will now be presen-
ted, pinpointing the optimal gas-turbine exhaust-gas temperature and corresponding minimum
compressor-inlet temperature, coupled with the optimal compressor pressure ratio (CPR), for
achieving maximum closed-combined-cycle thermal efficiency.
The isentropic compressor or turbine efficiency is usually used as a measure of the quality of
the aerodynamic design (or as a measure of friction losses) in an adiabatic machine. However
, it has a serious disadvantage, because its value is a function of pressure ratio and losses.
To avoid the influence of pressure ratio on the isentropic efficiency, the limiting value of the
isentropic efficiency for a given polytropic process can be used as pressure ratio approaches
unity. In steam turbines this type of efficiency is usually known as the "small-stage" efficiency,
while in gas-turbine expanders and compressors it is known as the "polytropic" efficiency. In
removing the effect of pressure ratio, the polytropic efficiency enables machines of a different
pressure ratio to be compared with validity. Therefore, in the mathematical model of this
invention only polytropic efficiencies are used in the thermodynamic equations.
There are two basic thermodynamic equations to apply in calculating compression and expan-
sion works and corresponding temperature changes. The first equation gives the relationship
of the temperature- and pressure-ratios of both compression and expansion processes, based
on the polytropic (small-stage) efficiencies of compressors and gas turbines:
(kC - 1) R
T2 = (P2)kc * ~PC = (P2~ cpc * IJPC
T, P1
...20
T4 P3 k * I~PE P3 ( c ) ~PE
T3 P4 P4
where: UT1" and UT2" are absolute actual temperatures before and after compression, UT3"
and UT4" are absolute actual temperatures before and after expansion, Up1" and Up2" are ab-
solute pressures before and after compression, Up3" and Up4' are absolute pressures before
and after expansion, UR" is the gas constant, "overlined C sub pC" is the mean constant-pres-
sure specific heat of the gas during compression, "overlined C sub pE" is the mean constant-
-pressure specific heat of the gas during expansion, "k sub C" is the average ratio of the spe-
cific heats (Cp/Cv) of the gas during compression, "k sub E" is the average ratio of the specific
heats of the gas during expansion, ",u sub PC" is the polytropic efficiency of the compressor
and "I~ sub PE" is the polytropic efficiency of the gas turbine.
The second formula deals with the change in enthalpy of the gas that is used as a working
fluid. Thus, combining this formula with the first one, the works of compression and expansion,
respectively, can be determined as:
(kC-1)
WC = (H2 ~ H,) = m * CPC * (T2 - T1) = m * CPC * T1 * [( P ) 1 ]
E ( 3 4) m CPE * (T3 - T4) = m * CPE * T3 * [1 _ (_) kE ]
' f~~
,~
... 21
where: '~ sub C" is the compression work (or power) required, "W sub E" is the expansion
work (or power) obtained, "(H2-H1 )" is the corresponding enthalpy change (rise) of the com-
pressed gas, "(H3-H4)" is the corresponding enthalpy change (drop) of the expanded gas, Dm"
is the mass (or mass flow-rate) of the gas, "(T2-T1 )" is the polytropic (real) temperature chan-
ge (rise) of the compressed gas and "(T3-T4)" is the polytropic (real) temperature change
(drop) of the expanded gas. All other designations have the same meanings as they do in the
previous formula.
Temperature designated UT4" is the actual absolute turbine-exhaust-gas temperature and also
the actual absolute temperature at the HRSG-inlet, which is the most important parameter for
obtaining maximum closed-combined-cycle thermal efficiency, according to the teachings of
this invention. If calculation of the compression/expansion works and thermal efficiency is to
be done for a closed-combined cycle with one (single-reheat) or two (dual-reheat) reheats,
then only the total expansion ratio has to be divided between two or three gas turbines, res-
pectively, and the corresponding actual turbine-exhaust-gas absolute temperature would be
designated as UT6" orUT8'', respectively.
Inevitable pressure parasitic losses are assumed to be 5% in the gas-heater and also through
the heat recovery steam generator, or 2.5% in each eventual gas-reheater. This can be math-
ematically described as: P3 = 0.95 * P2 and P1 = 0.95 * P4 (or P6), or P5 = 0.975 * P4.
In this mathematical model, constant-pressure specific heat of the gas is considered variable
with temperature (this is a "semi-perfect" gas model). Consequently, constant-pressure speci-
fic heats of all gases considered are given as potential functions of the gas temperature.
However, in practice the use of polytropic efficiencies does not eliminate all influence of pres-
sure ratio from efficiency. For example, the polytropic efficiencies of both compressor and ex-
pander of 0.9 (90%) can easily be attained at pressure ratios of up to 10 in the larger machi-
... 22
nes, but this polytropic efficiencies are unrealistically high for the higher pressure ratios. The-
refore, in this mathematical model the compressor and gas-turbine polytropic efficiencies are
assumed to be simple linear functions of the compression pressure ratio, as given in the follo-
wing relations:
og [(P2/P~ 6] 1~ = 0864 _ [Cp2lp1)-16]
As it was mentioned earlier, part of the compressor air is usually branched off in order to per-
form the cooling of the gas-turbine blading, because of extremely high gas-turbine-inlet tem-
perature T3. This is also a significant factor that diminishes the cycle thermal efficiency and,
consequently, it should be taken into account. This was done by assuming that the main detri-
mental effect due to turbine cooling is the associated reduction of hot-gas-steam temperature.
To estimate the equivalent-reduced gas-turbine-inlet temperature due to cooling, a careful
analysis was done of the available data for an existing Siemens Model V 84.3 advanced-po-
wer-generation gas turbine and a Westinghouse 501 F open-combined cycle.
The Siemens Model V 84.3 advanced gas turbine is claimed to have 36.1 % simple-cycle ther-
mal efficiency and 54% open-combined-cycle thermal efficiency, at the firing (burner-outlet)
temperature of 2350~F (1560.78 K) and the compression pressure ratio of 16. The exhaust-
-gas temperature is 1022~F (823 K). With these data incorporated in the mathematical model
of this invention, assuming that ambient-air temperature is 59~F (288 K), the calculated gas-
-turbine-inlet equivalent-reduced temperature is 1474.23 K with corresponding cooling-air
fraction of 6% by mass of the compressor-air mass flow-rate, if simple-cycle thermal efficiency
is to be 36.1 % and open-combined-cycle thermal efficiency is to be 54%. For firing temperatu-
res different from 1560.78 K, a parabolic change of colling-air fraction was assumed. Thus for-
mulated mathematical model was successfully tested for the Westinghouse 501 F open-combi-
,~ .
~,
...23
ned cycle and it showed very good agreement with available data. For information, this cycleuses a gas turbine with burner-outlet temperature of 2460~F (1349~C) and rotor-inlet tempera-
ture of 2300~F (1260~C) and it achieves LHV-thermal efficiency in the range of 54% to 55% at
ISO conditions. Because of very good agreement with available data, such formulated mathe-
matical model was adapted for closed-combined-cycles calculations and has been used for
estimating of thermal efficiencies and specific-power outputs of the closed-combined cycles.
A representative steam-turbine cycle, included in the all closed-combined-cycle configurati-
ons disclosed in this invention, was considered to be a conventional superheated/reheated
high-pressure-steam cycle with: superheated/reheated steam temperature of 540~C (813 K or
1004~F), superheated-steam pressure of 175 bars (2538 psi), reheated-steam pressure of 40
bars (580 psi) and condensation temperature of 30~C (303 K or 86~F). The adiabatic expan-
sion efficiency of each part of the steam turbine (high-, intermediate- and low-pressure steam
turbines) is assumed to be 86%. With these assumptions, all steam and condensate thermo-
dynamic properties were taken or interpolated from the steam-water tables. Works required by
the condensate and feedwater pumps were also taken into account, assuming that isentropic
pumping efficiency is 65%.
After the introduction of applied mathematical model, the explanation of FIG. 6 can be given.
FIG. 6 represents a typical change-of-temperature-versus-length graph in a heat exchanger, a
heat recovery steam generator (HRSG) in this case. It can be noted that the temperature
change of the exhaust gas (the hotter fluid in the graph) is represented by a straight line; this
assumption is very close to an actual temperature distribution in an HRSG. The minimum tem-
perature difference of the two fluids is designated by Udelta T sub min" in the graph and it is
assumed to be 3 K (5.4~F),the assumption being based on the estimation of modern HRSGs
from available data. From the graph, it can be observed that three different types of lines are
depicted, designating 3 corresponding exhaust-gas temperatures. A solid line designates the
'" ~
... 24
exhaust-gas temperature of 1067 K (1461 ~F), which is determined to be an optimum exhaust-
-gas temperature for obtaining maximum air-closed-combined-cycle thermal efficiency, provi-
ded the exhaust gas can be cooled down to the minimum HRSG-outlet (or compressor-inlet)
temperature of 310 K (98.6~F), according to this invention. Taking into account that an assu-
med condensation temperature is 303 K (86~F), this minimum HRSG-outlet temperature is
achievable, with corresponding temperature difference at the HRSG-outlet of 7 K (12.6~F).
A dotted line designates the exhaust-gas temperature of 967 K (1281~F), which is 100 K
(180~F) lower than the optimum one. It can be noted that, for the same minimum temperature
difference of the two fluids, a higher HRSG-outlet (or compressor-inlet) temperature appears,
which c~uses a decrease in closed-combined-cycle thermal efficiency.
On the other hand, a dashed line designates the exhaust-gas temperature of 1167 K (1641 ~F)
, which is 100 K (180~F) higher than the optimum one. Now it can be noted that, for the same
achievable HRSG-outleVcompressor-inlet temperature of 310 K (98.6~F), a larger "minimum"
temperature difference of the two fluids will appear from the condition that the exhaust-gas
temperature exceeds the optimum one. Again, this will cause a similar decrease in closed-
~ombined-cycle thermal efficiency.
Consequently, a key fact that has been taken advantage of in this invention, is the following:
the higher the turbine-exhaust-gas temperature at the inlet of the heat recoverv steam
generator (HRSG), the lower the compressor-inlet-gas temperature will be, approaching
to an optimal exhaust-gas temperature for a certain achievable compressor-inlet tempe-
rature, assuming linear temperature change of the gas across the HRSG. According to
the invention, this statement is completely true for closed-combined cycles with one or more
gas-reheats, whereas it is a little bit different for combined cycles without gas-reheats. In the
latter case, the compressor-inlet temperature does not necessarily have to be as low as possi-
... 25
ble, so that maximum cycle efficiency can be obtained for a certain combination of the optimalturbine-exhaust-gas and compressor-inlet temperatures, for an optimal compression pressure
ratio (CPR).
An illustration and an evidence of this conclusion are given in FIGS. 7, 8, 9, 10, 11, 12, 13 and
14, where thermal efficiency of different closed-combined-cycle configurations (with different
working fluids: dry air, helium and hydrogen) is plotted versus net specific power, for 3 differ-
ent ",a~i",um cycle temperatures (1475 K, 1550 K and 1625 K). Each figure relates to a parti-
cular closed-combined-cycle configuration with a certain working gas. All figures are clearly
named and briefly explained in the topic "Brief Description of the Drawings". It can be noted
that in all closed-combined cycles with one (single) or two (dual) gas-reheats, an optimal gas-
-turbine exhaust-gas temperature for maximum cycle thermal efficiency is close to 1070 K, al-
most regardless from the working fluid. For closed-combined cycles without gas-reheats, an
optimal CPR determines maximum cycle thermal efficiency and also optimal exhaust-gas tem-
perature, which is lower than 1070 K, but still significantly higher than usually used ones.
From the mathematical model applied with all assumptions and restrictions, it appears that the
best configuration of the closed-combined cycle is the one with single gas-reheat. If helium is
used as a working gas in this configuration, then, according to FIG. 10, the maximum cycle
thermal efficiency of 60% can be achieved with gas-turbine-inlet temperature of 1625 K. The
resulting net specific power is 3736 kJ/kg, of which 1627 kJ/kg (or 43.5%) is produced by the
bottoming steam-cycle. With assumed helium mass flow-rate of 475 kg/s (which is the mass
flow-rate of the Siemens Model V 84.3 open-simple-cycle gas turbine), the total electrical po-
wer of about 1770 MW can be achieved with such a closed-combined-cycle power-plant with
helium as a working gas, of which power -770 MW would yield the bottoming steam-cycle.
With hydrogen this power output could be obtained with approximately 3-times smaller mass
flow-rate, at the similar cycle thermal efficiency. This means that with a helium or hydrogen
... 26
single-gas-reheat closed-combined cycle it is possible to build a new, or repower an existing
steam-turbine cycle of any size. Of course, this is also true for helium or hydrogen two- or
multi-gas-reheats closed-combined-cycle configurations, but in these cases cycle thermal
efficiency will decrease due to very detrimental influence of an increased cooling-gas fraction,
which increases not only with the gas-turbine-inlet temperature, but also with the number of
gas-reheats, that is, with the number of gas turbines in a combined cycle.
An optimistic future prediction for closed-combined-cycles application is given in FIGS. 15,16,
17 and 18, where closed-combined-cycle thermal efficiency is plotted versus CPR for helium
and hydrogen single- and dual-gas-reheats configurations with optimal exhaust-gas tempera-
ture of 1080 K, assuming that closed-loop steam-cooling system is used for cooling of all the
gas-turbines's airfoils. It can be noted that very high closed-combined-cycle thermal efficiency
could be achieved: more than 63% for single-gas-reheat configurations and 64% to 65% for
dual-gas-reheats configurations. Connecting these numbers with very large net specific power
outputs obtainable with such helium (-4000 kJ/kg) and hydrogen (-12000 kJ/kg) closed-com-
bined-cycle configurations, it is very easy to understand a huge importance of the closed-com-
bined cycle with optimal exhaust-gas temperature and closed-loop steam-cooling system, as
proposed in this invention, in power-producing engineering.
It is interesting to point out that in the proposed closed-combined-cycle configurations much
larger part of the total cycle power output is produced by the steam-turbine bottoming cycle
than in typical open-cycle configurations, this part sometimes being greater than 50% of the
total closed-cycle output. For example, the percentage of the steam-turbine-cycle power out-
put in the total closed-combined-cycle power output is from 48% to 52% for a combined cycle
without dry-air-reheats (49% to 54% for helium-cycle) and from 43.5% to 47% for a combined
cycle with one or two gas-reheats (dry-air-, helium- or hydrogen-cycle), respectively, over the
considered range of gas-turbine-inlet temperatures (1625 K to 1475 K).
~ r s~ Y~ 27
One of the greatest advantages of the proposed closed combined cycle is its independence
from the type of fuel used, so that literally any type of fuel can be burnt in its furnace. This is
of particular importance with respect to coal, as the most widespread fossil fuel in the world
and, also, solid or liquid waste materials considered for incineration. So, with a closed-combi-
ned-cycle power-plant in accordance with this invention, it is possible to convert chemical
energy of fuel, particularly coal, into mechanical energy at considerably higher efficiency than
with previous combined-cycle systems and there is no longer need for coal gasification, as in
the case of open combined cycles. With an assumption that the closed-cycle furnace heat-ex-
change effectiveness is 90%, which is approximately the highest limit for a conventional steam
boiler, the total efficiency of the conversion of coal chemical energy into mechanical (electri-
cal) energy amounts to 54% for a closed-combined single-gas-reheat cycle, which is a much
higher value than that of a conventional open-combined cycle with coal gasification (~43%). If
closed-loop steam cooling of the gas-turbine airfoils is used, then it is possible to convert 58%
of coal chemical energy into mechanical energy with a closed-combined dual-gas-reheat cyc-
le. This percentage could be even higher, provided that higher gas-furnace effectiveness may
be achieved, since the mean heat-exchange temperature of a gas furnace is considerably
(approximately twice) higher than that of a conventional steam boiler. Such a mean-heat-ex-
change-temperature increase is caused by extremely high gas temperatures needed. Also,
because of fairly high compressor-outlet temperatures, the combustion air can be preheated
to higher temperatures in the gas furnace, resulting in a further increase of its effectiveness.
For illustration of the fact that very high gas-furnace efficiencies can be achieved today, the
following information will be given: the recuperator (exhaust-gas/compressed-air heat exchan-
ger) of an advanced open-gas-turbine system has been designed, which can achieve an effe-
ctiveness of 94.2%, with a combined (both sides) pressure drop of no more than 3.45% at an
exhaust-gas mass flow-rate of about 520 kg/s (1146 Ib/s). For higher temperatures of exhaust
gas a new material has been developed, which has 1 300~F (704~C) temperature capability.
;~
... 28
Material-development work is currently being done at 1 500~F to 1 800~F (81 5~C to 980~C).
The principles, preferred embodiments and modes of operation of the present invention have
been described in the foregoing specification. Since the invention is capable of further modifi-
cations, this application is intended to cover any eventual variations, uses or adaptations of
the invention, which can be done following the general principles of the invention and included
within the spirit and scope of the appended claims.
~,;
