U.S. patent application number 09/883551 was filed with the patent office on 2001-12-06 for turbo-machines.
Invention is credited to Griffiths, Kenneth F..
Application Number | 20010047648 09/883551 |
Document ID | / |
Family ID | 46257797 |
Filed Date | 2001-12-06 |
United States Patent
Application |
20010047648 |
Kind Code |
A1 |
Griffiths, Kenneth F. |
December 6, 2001 |
Turbo-machines
Abstract
A gas turbo-machine and method of designing and constructing
such machine includes preselecting specific operating conditions
for the gas turbo-machine, and constructing a master stage as a
model to have a given design and geometric shape which results in
substantially the optimum efficiency during operation of the master
stage at the preselected operating conditions. At least one
additional stage is then added to the master stage which is
substantially identical to the master stage in geometric shape and
design, but in which the linear dimensions of the additional stage
differ from those of the master stage in accordance with the
formula 1 L T = D T 3 where L is the ratio of the linear dimensions
of the additional stage to the master stage and D is the gas
density ratio of the master stage. The turbo-machines of the
present invention may be either axial or radial flow, compressors
or gas turbines, and isothermal, adiabatic or combinations thereof
in operation, there may or may not be tributary flow to or from one
or more of the stages and the turbo-machines may be utilized in the
generation of electrical power.
Inventors: |
Griffiths, Kenneth F.; (Fort
Atkinson, WI) |
Correspondence
Address: |
COOK, ALEX, MCFARRON, MANZO, CUMMINGS & MEHLER LTD
SUITE 2850
200 WEST ADAMS STREET
CHICAGO
IL
60606
US
|
Family ID: |
46257797 |
Appl. No.: |
09/883551 |
Filed: |
June 18, 2001 |
Related U.S. Patent Documents
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Application
Number |
Filing Date |
Patent Number |
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09883551 |
Jun 18, 2001 |
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09527994 |
Mar 17, 2000 |
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6260349 |
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Current U.S.
Class: |
60/774 ;
60/735 |
Current CPC
Class: |
Y10T 29/4932 20150115;
F01D 1/10 20130101; F02C 6/003 20130101; F02C 6/203 20130101; F02C
7/143 20130101; F02C 7/185 20130101 |
Class at
Publication: |
60/39.04 ;
60/735 |
International
Class: |
F02C 006/00; F02C
003/16 |
Claims
What is claimed is:
1. A multistage gas turbo-machine comprising a first stage and a
subsequent stage of differing sizes, each stage having
turbo-machine blades which are contacted by the gas, an inlet in
each stage for introducing the gas to the turbo-machine blades in
the stage, a discharge from each stage for discharging the gas from
the turbo-machine blades in the stage, and the discharge from one
stage communicating with the inlet of the other stage; and wherein
said first and subsequent stages are substantially identical to
each other in design and geometric shape, but in which the linear
dimensions of said subsequent stage differ from those of said first
stage substantially in accordance with the formula 82 L T = D T 3
where L.sub.T is the ratio of the linear dimensions of the
subsequent stage to the first stage when said subsequent stage is
downstream of said first stage, and D.sub.T is the gas density
ratio of said subsequent stage, and 83 D T = ( M A + M B tot M A )
( P I / P O T I / T O ) n - 1 where: M.sub.A=molar volume flow rate
to intake of stage 1, moles/sec; M.sub.B tot=total tributary volume
flow rate generated in or between all preceding stages, moles/sec;
P.sub.I=absolute pressure of gas entering stage in question;
P.sub.O=absolute pressure of gas leaving stage in question;
T.sub.I=absolute temperature of gas entering stage in question;
T.sub.O=absolute temperature of gas leaving stage in question; and
n=number of the stage in question.
2. The gas turbo-machine of claim 1, including a power transmission
shaft, and wherein at least some of said turbo-machine blades are
coupled to said shaft to rotate with said shaft, and said shaft and
the rotating turbo-machine blades of said first and subsequent
stages rotate at the same speed.
3. The gas turbo-machine of claim 2, wherein said machine is either
an axial flow or a radial flow gas turbo-machine.
4. The gas turbo-machine of claim 3, wherein said machine is a
compressor, and the linear dimensions of said subsequent stage are
smaller than the linear dimensions of said first stage
substantially in accordance with said formula.
5. The gas turbo-machine of claim 4, wherein said first and
subsequent stages are substantially isothermal.
6. The gas turbo-machine of claim 5, wherein at least said first
stage also includes stator blades, and said stator blades include
an inlet and outlet for passing a coolant through said blades to
cool the gas to said substantially isothermal conditions before the
gas is discharged from said first stage.
7. The gas turbo-machine of claim 4, wherein at least said first
stage is substantially adiabatic.
8. The gas turbo-machine of claim 7, including an intercooler
between said first stage and a next stage to cool the gas
discharged from said first stage before the gas enters the inlet of
the next stage.
9. The gas turbo-machine of claim 8, wherein said intercooler cools
the gas to substantially the same temperature as the gas introduced
to the inlet of said first stage.
10. The gas turbo-machine of claim 3, wherein said machine is a gas
turbine, and the linear dimensions of said subsequent stage are
larger than the linear dimensions of said first stage substantially
in accordance with said formula.
11. The gas turbo-machine of claim 10, wherein said first and
subsequent stages are substantially isothermal.
12. The gas turbo-machine of claim 11, wherein at least said first
stage also includes a fuel injector which injects fuel into said
first stage to heat the gas to said substantially isothermal
conditions before it is discharged from said first stage.
13. The gas turbo-machine of claim 10, wherein at least said first
stage is substantially adiabatic.
14. The gas turbo-machine of claim 13, including a combustor
between said first stage and a next stage which heats the gas
discharged from said first stage before the gas enters the inlet of
the next stage.
15. The gas turbo-machine of claim 14, wherein said combustor heats
the gas to substantially the same temperature as the gas introduced
to the inlet of said first stage.
16. The gas turbo-machine of claim 1, wherein said machine is a
compressor, and the linear dimensions of said subsequent stage are
smaller than the linear dimensions of said first stage
substantially in accordance with said formula.
17. The gas turbo-machine of claim 16, wherein said first and
subsequent stages are substantially isothermal.
18. The gas turbo-machine of claim 17, wherein at least said first
stage also includes stator blades, and said stator blades include
an inlet and outlet for passing a coolant through said blades to
cool the gas to said substantially isothermal conditions before the
gas is discharged from said first stage.
19. The gas turbo-machine of claim 16, wherein at least said first
stage is substantially adiabatic.
20. The gas turbo-machine of claim 19, including an intercooler
between said first stage and a next stage to cool the gas
discharged from said first stage before the gas enters the inlet of
the next stage.
21. The gas turbo-machine of claim 20, wherein said intercooler
cools the gas to substantially the same temperature as the gas
introduced to the inlet of said first stage.
22. The gas turbo-machine of claim 1, wherein said machine is a gas
turbine, and the linear dimensions of said subsequent stage are
larger than the linear dimensions of said first stage substantially
in accordance with said formula.
23. The gas turbo-machine of claim 22, wherein said first and
subsequent stages are substantially isothermal.
24. The gas turbo-machine of claim 23, wherein at least said first
stage also includes a fuel injector which injects fuel into said
first stage to heat the gas to said substantially isothermal
conditions before it is discharged from said first stage.
25. The gas turbo-machine of claim 22, wherein at least said first
stage is substantially adiabatic.
26. The gas turbo-machine of claim 25, including a combustor
between said first stage and a next stage to heat the gas
discharged from said first stage before the gas enters the inlet of
the next stage.
27. The gas turbo-machine of claim 26, wherein said combustor heats
the gas to substantially the same temperature as the gas introduced
to the inlet of said first stage.
28. The gas turbo-machine of claim 2, including a generator for
generating electrical power, said power transmission shaft
mechanically coupling said turbine blades with said generator.
29. The gas turbo-machine of claim 28, including a compressor and a
gas turbine, at least one of said compressor and said gas turbine
including said first and subsequent stages, and the gas from said
compressor is discharged to said gas turbine; and a heat exchanger
positioned between said compressor and said gas turbine and in the
discharge from said gas turbine to heat the gas being discharged
from said compressor before it is introduced to said gas turbine
with the heat from the gas which is discharged from said gas
turbine.
30. The gas turbo-machine of claim 29, wherein both said compressor
and said gas turbine each contain said first and subsequent
stages.
31. The gas turbo-machine of claim 29, wherein said compressor
contains said first and subsequent stages, and the linear
dimensions of said subsequent stage are smaller than the linear
dimensions of said first stage substantially in accordance with
said formula.
32. The gas turbo-machine of claim 31, wherein said first and
subsequent stages are substantially isothermal.
33. The gas turbo-machine of claim 32, including means for
introducing water to said first stage from below the thermocline of
a large body of water, wherein said first stage also includes
stator blades, and said stator blades include an inlet and outlet
for passing said water through said blades to cool the gas to said
substantially isothermal conditions before the gas is discharged
from said first stage.
34. The gas turbo-machine of claim 31, wherein at least said first
stage is substantially adiabatic.
35. The gas turbo-machine of claim 34, including an intercooler
between said first stage and a next stage to cool the gas
discharged from said first stage before the gas enters the inlet of
the next stage.
36. The gas turbo-machine of claim 35, wherein said intercooler
cools the gas to substantially the same temperature as the gas
introduced to the inlet of said first stage.
37. The gas turbo-machine of claim 36, including means for
introducing water to said intercooler from below the thermocline of
a large body of water to cool said gas.
38. The gas turbo-machine of claim 29, wherein said gas turbine
contains said first and subsequent stages, and the linear
dimensions of said subsequent stage are larger than the linear
dimensions of said first stage in accordance with said formula.
39. The gas turbo-machine of claim 38, wherein said first and
subsequent stages are substantially isothermal.
40. The gas turbo-machine of claim 39, wherein at least said first
stage also includes a fuel injector which injects fuel into said
first stage to heat the gas to said substantially isothermal
conditions before it is discharged from said first stage.
41. The gas turbo-machine of claim 38, wherein at least said first
stage is substantially adiabatic.
42. The gas turbo-machine of claim 41, including a combustor
between said first stage and a next stage which heats the gas
discharged from said first stage before the gas enters the inlet of
the next stage.
43. The gas turbo-machine of claim 42, wherein said combustor heats
the gas to substantially the same temperature as the gas introduced
to the inlet of said first stage.
44. The gas turbo-machine of claim 1, wherein when no tributary
volume flow rate is generated in or between a stage, 84 ( M A + M B
tot M A ) is 1 and D T = ( P I / P O T I / T O ) n - 1
45. A method of designing and constructing a multistage gas
turbo-machine comprising the steps of: preselecting the operating
conditions for the gas turbo-machine of gas pressure ratio, gas
intake temperature and gas air flow rate; constructing a master
stage to have a given design and geometric shape which results in
substantially the optimum efficiency during operation of said
master stage under the preselected operating conditions; and
constructing at least one additional subsequent stage of said
multistage gas turbo-machine, said additional subsequent stage
being substantially identical to said master stage in geometric
shape and design, but in which the linear dimensions of said
additional subsequent stage differ from those of said master stage
substantially in accordance with the formula 85 L T = D T 3 where L
is the ratio of the linear dimensions of the additional subsequent
stage to the master stage when the additional subsequent stage is
downstream of the master stage, and D.sub.T is the gas density
ratio of the additional subsequent stage, and 86 D T = ( M A + M B
tot M A ) ( P I / P O T I / T O ) n - 1 where: M.sub.A=molar volume
flow rate to intake of stage 1, moles/sec; M.sub.B tot=total
tributary volume added to flow rate generated in or between all
preceding stages, moles/sec; P.sub.I=absolute pressure of gas
entering stage in question; P.sub.O=absolute pressure of gas
leaving stage in question; T.sub.I=absolute temperature of gas
entering stage in question; T.sub.O=absolute temperature of gas
leaving stage in question; and n=number of the stage in
question.
46. The method of claim 45 wherein each of said stages include
turbo-machine rotor blades, and said turbo-machine rotor blades are
coupled to a power transmission shaft to rotate with said shaft and
so that the rotating turbo-machine rotor blades of the master and
additional subsequent stages rotate at the same speed.
47. The method of claim 45, wherein the gas turbo-machine is a
compressor, and wherein a coolant supply is provided to cool the
gas before it enters the additional subsequent stage.
48. The method of claim 47, wherein said coolant supply is
constructed to cool the gas which is discharged from a stage to
substantially the same temperature as the gas entered that stage,
whereby said stages will operate under substantially isothermal
conditions.
49. The method of claim 47, wherein said coolant supply is
constructed to cool the gas after it is discharged from the master
stage, but before the gas enters the additional subsequent stage to
substantially the same temperature as the gas introduced to the
master stage.
50. The method of claim 45, wherein the gas turbo-machine is a gas
turbine, and a heat supply is provided to heat the gas before it
enters the additional subsequent stage.
51. The method of claim 50, wherein said heat supply is constructed
to heat the gas which is discharged from a stage to substantially
the same temperature as the gas entered that stage, whereby said
stages will operate under substantially isothermal conditions.
52. The method of claim 50, wherein said heat supply is constructed
and arranged to heat the gas after it is discharged from the master
stage, but before the gas enters the additional subsequent stage to
substantially the same temperature as the gas introduced to the
master stage.
53. The gas turbo-machine of claim 1, wherein when no tributary
volume flow rate is generated in or between a stage, 87 ( M A + M B
tot M A ) is 1 and D T = ( P I / P O T I / T O ) n - 1
Description
BACKGROUND AND SUMMARY OF THE INVENTION
[0001] The present invention is directed to turbo-machines and,
more particularly, to multistage axial or radial gas flow
compressors and turbines and systems employing such
turbo-machines.
[0002] It is known that the efficiency of turbo-machines, such as
compressors and gas turbines, may be substantially improved by
operation in a manner which approaches isothermal conditions. This
essentially means that the temperature of the gas as it moves
between successive stages of the turbo-machine is adjusted so that
the inlet temperature of the gas at each successive stage is
maintained at about the same temperature as at the inlet of the
preceding stage. This is in contrast to adiabatic operation in
which the temperature of the gas changes between the successive
stages due to the compression or expansion of the gas as it moves
through each successive stage of the turbo-machine.
[0003] Maintenance of a constant temperature at the inlet of each
successive stage may be accomplished in several different ways. In
a purely isothermal gas turbine, fuel injectors and temperature
sensors may be positioned in each stage so that the correct amount
of fuel is injected into and burned in each stage as is needed to
ensure that the temperature of the gas in the gas turbine is
re-elevated to substantially the temperature at which it entered
that stage prior to discharge from the stage and introduction to
the next succeeding stage. This is shown for example in U.S. Pat.
No. 4,197,700 (Jahnig). In a purely isothermal compressor, a
coolant may be introduced into each stage, for example through the
stator blades of an axial compressor, to reduce the temperature of
the gas to substantially the same temperature at which it was
introduced to that stage to ensure that the temperature of the gas
which is discharged from the stage and introduced to the next stage
is at substantially the same temperature. Combustion chambers or
intercoolers have also been employed between stages to add or
remove heat and alter the gas temperature so that the gas entering
each of the respective stages is at substantially the same
temperature.
[0004] Substantial improvements in efficiency may also be achieved
in particular in compressors through the use of relatively low
temperature coolants, such as sea water which is taken from below
the thermocline. Such sea water will typically be about 40.degree.
F. which is sufficient to maintain a temperature of about
45.degree. F. to the intake of each stage of an isothermal
compressor.
[0005] It would also be desirable to design, for example, the first
stage of the turbo-machine to achieve the maximum efficiency from a
design standpoint when the turbo-machine is in normal operation.
Normal operation means that each stage would have a given shaft
speed, pressure ratio, temperature ratio, gas density ratio, and
the type of operation in each stage would be the same, e.g.
isothermal, adiabatic, etc. This optimum efficiency stage could
then act as a master stage which would serve as a model for the
construction of each of the subsequent stages. In the present
invention a formula has been discovered for the sizing of each
subsequent stage once an optimum efficiency master stage has been
designed which will maximize the optimum efficiency of each
subsequent stage so that it has substantially the same optimum
efficiency as the optimum efficiency master stage.
[0006] It has also been discovered that the sizing formula of the
present invention is applicable to all turbo-machines whether they
are purely isothermal in operation, purely adiabatic in operation,
or a combination of adiabatic/isothermal operation as in
turbo-machines employing intercoolers or intercombustion chambers
between stages to adjust the temperature of the gas to a given
selected temperature prior to introduction of the gas to the next
successive stage. And, it has been discovered that the sizing
formula of the present invention is also equally applicable to
either axial flow or radial flow turbo-machines, and to a wide
range of types of turbo-machines including compressors, gas
turbines and gas expanders. Significantly, the sizing formula of
the present invention may be utilized in the sizing of the
turbo-machine stages whether or not a tributary gas flow is
introduced to or removed from one or more of the stages. Such
tributary flow may be introduced for example to each stage in the
form of fuel to provide for isothermal operation.
[0007] Gas expanders are quite similar in construction to gas
turbines, but each has a somewhat different emphasis and purpose.
In both gas turbines and gas expanders the gas expands as it moves
through the several successive stages. However, gas turbines
generally have the purpose of generating drive shaft power, for
example to power an electrical generator, whereas gas expanders
have the principal function of permitting a controlled expansion of
gases for the purpose of cooling the gas. Because of the similarity
of construction of gas turbines and expanders, the term "gas
turbine" as employed hereinafter will include both gas turbines as
well as gas expanders, unless otherwise stated.
[0008] In one principal aspect of the present invention, a
multistage gas turbo-machine includes a first stage and a
subsequent stage of differing sizes. Each stage has turbo-machine
blades which are contacted by the gas, an inlet in each stage for
introducing the gas to the turbo-machine blades in the stage, a
discharge from each stage for discharging the gas from the
turbo-machine blades in the stage, and the discharge from one stage
communicates with the inlet of the other stage. The first and
subsequent stages are substantially identical to each other in
design and geometric shape, but the linear dimensions of the
subsequent stage differ from those of the first stage substantially
in accordance with the formula 2 L T = D T 3
[0009] where L.sub.T is the ratio of the linear dimensions of the
subsequent stage to the first stage and D is the gas density ratio
of the first stage, and 3 D T = ( M A + M B tot M A ) ( P I / P O T
I / T O ) n - 1
[0010] where:
[0011] M.sub.A=molar volume flow rate to intake of stage 1,
moles/sec;
[0012] M.sub.B tot=total tributary volume added to or between all
preceding stages, moles/sec;
[0013] P.sub.I=absolute pressure of gas entering stage in
question;
[0014] P.sub.O=absolute pressure of gas leaving stage in
question;
[0015] T.sub.I=absolute temperature of gas entering stage in
question;
[0016] T.sub.O=absolute temperature of gas leaving stage in
question; and
[0017] n=number of the stage in question.
[0018] In another principal aspect of the present invention, the
gas turbo-machine includes a power transmission shaft, and at least
some of the turbine blades are coupled to the shaft to rotate with
the shaft, and the shaft and the rotating turbine blades of the
first and subsequent stages rotate at the same speed.
[0019] In still another principal aspect of the present invention,
the gas turbo-machine is either an axial flow or a radial flow gas
turbo-machine.
[0020] In still another principal aspect of the present invention,
the gas turbo-machine is a compressor, and the linear dimensions of
the subsequent stage are smaller than the linear dimensions of the
first stage substantially in accordance with the formula.
[0021] In still another principal aspect of the present invention,
the first and subsequent stages of the compressor are substantially
isothermal.
[0022] In still another principal aspect of the present invention,
the first stage of the compressor also includes stator blades, and
the stator blades include an inlet and outlet for passing a coolant
through the blades to cool the gas to the substantially isothermal
temperature before the gas is discharged from the first stage.
[0023] In still another principal aspect of the present invention,
at least the first stage of the compressor is substantially
adiabatic.
[0024] In still another principal aspect of the present invention,
the compressor includes an intercooler between the first stage and
a next stage to cool the gas discharged from the first stage before
the gas enters the inlet of the next stage.
[0025] In still another principal aspect of the present invention,
the intercooler cools the gas to substantially the same temperature
as the gas introduced to the inlet of the first stage.
[0026] In still another principal aspect of the present invention,
the gas turbo-machine is a gas turbine, and the linear dimensions
of the subsequent stage are larger than the linear dimensions of
the first stage substantially in accordance with the formula.
[0027] In still another principal aspect of the present invention,
the first and subsequent stages of the gas turbine are
substantially isothermal.
[0028] In still another principal aspect of the present invention,
the first stage of the gas turbine also includes a fuel injector
which injects fuel into the first stage to heat the gas to the
substantially isothermal temperature before it is discharged from
the first stage.
[0029] In still another principal aspect of the present invention,
at least the first stage of the gas turbine is substantially
adiabatic.
[0030] In still another principal aspect of the present invention,
the gas turbine includes a combustor between the first stage and
the next stage which heats the gas discharged from the first stage
before the gas enters the inlet of the second stage.
[0031] In still another principal aspect of the present invention,
the combustor heats the gas to substantially the same temperature
as the gas introduced to the inlet of the first stage.
[0032] In still another principal aspect of the present invention,
the gas turbo-machine includes a generator for generating
electrical power, and the aforementioned power transmission shaft
mechanically couples the turbine blades with the generator.
[0033] In still another principal aspect of the present invention,
the gas turbo-machine with the generator includes a compressor and
a gas turbine, one or both of which includes the aforementioned
first and subsequent stages. The gas from the compressor is
discharged to the gas turbine, and a heat exchanger (regenerator)
is positioned between the compressor and the gas turbine. The
discharge from the gas turbine is used to heat the gas being
discharged from the compressor before it is introduced to the gas
turbine with the heat content of the gas which is discharged from
the gas turbine.
[0034] In still another principal aspect of the present invention,
water is introduced to the first stage of the compressor from below
the thermocline of a large body of water, the first stage also
includes stator blades, and the stator blades include an inlet and
outlet for passing the water through the blades to cool the gas to
the substantially isothermal temperature before the gas is
discharged from the first stage.
[0035] In still another principal aspect of the present invention,
a method of designing and constructing a multistage gas
turbo-machine comprises preselecting the operating conditions for
the gas turbo-machine of gas pressure ratio, gas intake temperature
and gas flow rate. A master stage is constructed to have a given
design and geometric shape which results in substantially the
optimum efficiency during operation of the master stage under the
preselected operating conditions. At least one additional
subsequent stage is then constructed which is substantially
identical to the master stage in geometric shape and design, but in
which the linear dimensions of the additional stage differ from
those of the master stage substantially in accordance with the
aforementioned formula.
[0036] In still another principal aspect of the present invention,
when no tributary volume flow rate is generated in or between a
stage, 4 ( M A + M B tot M A ) is 1 and D T = ( P I / P O T I / T O
) n - 1
[0037] These and other objects, features and advantages of the
present invention will be more clearly understood through a
consideration of the following detailed description.
BRIEF DESCRIPTION OF THE DRAWINGS
[0038] In the course of this description, reference will frequently
be made to the attached drawings in which:
[0039] FIG. 1 is a schematic depiction of an electric power
generation system utilizing one or more turbo-machines of the
present invention, and incorporating intercoolers and/or
intercombustors between the stages of the turbo-machines to adjust
the temperatures there between;
[0040] FIG. 2 is a broken, partial, cross-sectioned elevation view
of a schematic depiction of a portion of a preferred embodiment of
an isothermal axial compressor in accordance with the
invention;
[0041] FIG. 3 is a broken, partial, cross-sectioned elevation view
of a schematic depiction of a portion of a preferred embodiment of
an isothermal radial compressor in accordance with the
invention;
[0042] FIG. 4 is a broken, partial, cross-sectioned elevation view
of a schematic depiction of a portion of a preferred embodiment of
an isothermal axial turbine in accordance with the invention;
[0043] FIG. 5 is a broken, partial, cross-sectioned elevation view
of a schematic depiction of a portion of a preferred embodiment of
an isothermal radial turbine in accordance with the invention;
[0044] FIG. 6 is a broken, partial isometric view of a schematic
depiction of a typical rotor or stator blade of a turbo-machine;
and
[0045] FIG. 7 is a schematic depiction of a gas turbo-machine
having a tributary input to each of its stages.
DESCRIPTION OF THE PREFERRED EMBODIMENTS
[0046] With particular reference to FIG. 1, a schematic depiction
of a power generation system is shown which incorporates one or
more of the preferred turbo-machines, i.e. compressors and/or gas
turbines, constructed in accordance with the present invention. The
power generation system 10 as shown in FIG. 1, preferably includes
a single axial drive shaft 12 to which the rotors 14 of each of the
stages of a compressor 16 constructed in accordance with the
principles of the present invention are mechanically coupled to be
driven by the drive shaft 12. As seen in the drawings, the
compressor 16 is shown as comprising five stages 1-5. However, it
will be appreciated that the compressor 16 may consist of more or
less than five stages.
[0047] Also mechanically coupled to the drive shaft 12 are the
rotors 18 of a gas turbine 20. As such, the rotors 18 will power
the drive shaft 12 and the compressor 16, as well as an electrical
generator 22, which is also coupled to the drive shaft 12 to
generate electrical power.
[0048] The system shown in FIG. 1 comprises a plurality of
intercoolers C1-C5 through which the incoming air which is to be
compressed in the compressor 16 is circulated between each stage
1-5 of the compressor 16. The intercoolers C1-C5 preferably adjust
the temperature of the gas between each stage just before the gas
enters the next stage so that the gas is at substantially the same
temperature as the gas which entered the preceding stage.
[0049] The intercoolers C1-C5 are cooled by a suitable coolant
source which, as shown in FIG. 1, is sea water. The sea water is
preferably pumped from about 500-1500 feet below the surface and
below the thermocline TC, but a sufficient distance above the sea
floor F to prevent sediment and other bottom debris from being
drawn into the coolant system. The temperature beneath the
thermocline TC is typically quite low and may be on the order of
about 40.degree. F. This should insure that the temperature of the
gas passing through the compressor 16 and between its stages can be
as low as 45.degree. F. Such temperatures should result in a highly
efficient optimum system. Pumping of the sea water may be
accomplished by a suitable pumping station, such as ship S shown in
FIG. 1, which discharges the coolant through a main 24 to the
intercoolers C1-C5. The intercoolers C1-C5 are preferably arranged
in parallel to the main 24 so that the coolant to each intercooler
may be individually controlled by valves 25 as needed to maintain
the desired substantially uniform gas temperature at each stage.
The coolant is discharged from the intercoolers C1-C5 back to the
environment, for example to the surface SS of the sea as shown in
FIG. 1.
[0050] Also as shown in FIG. 1, interburners B1-B5 are positioned
in combustion chambers between each of the stages 1-5 of the
turbine 20. These burners receive the compressed air from the last
stage 5 of compressor 16 via a heat exchanger or regenerator 26.
The heat exchanger 26 preheats the incoming compressed air to the
turbine burner B1 with the heat which has been scavenged from the
hot gasses which are exhausted from the last stage 5 of the turbine
20. This preheating and scavenging also greatly optimizes and
improves the efficiency of the system, as well as cools the
discharged exhaust gases from the last stage 5 of the turbine 20 to
prepare those gases for discharge to the environment, such as
through an exhaust stack 28. The purpose of the interburners B1-B5
is to add energy to the gas passing between the stages to reheat
the gas just before it enters the next stage to a temperature which
is substantially the same temperature as the gas which entered the
preceding stage.
[0051] It will be appreciated that the power generation system
shown in FIG. 1 and thus far described in relation to that figure,
is effectively a hybrid adiabatic/isothermal system. That is a
system in which the temperature of the gas will rise between the
inlet and discharge from any given stage in the compressor, and
will fall between the inlet and discharge of any given stage in the
turbine as it would in a simple, conventional purely adiabatic
compressor and/or turbine. In such conventional adiabatic
turbo-machines the gas which is discharged from a stage at whatever
its temperature is after passing through that stage is simply
introduced to the inlet of the next succeeding stage. In such
conventional adiabatic system, the gas discharged from a stage and
introduced to the next succeeding stage is cooler in the case of a
turbine and hotter in the case of a compressor than when it was
introduced to the stage from which it is being discharged. However,
in the hybrid adiabatic/isothermal system shown in FIG. 1, the gas
discharged from each preceding stage is cooled by the intercoolers
C1-C5 in the case of the compressor and heated by the interburners
Bl-B5 in the case of the turbine to a temperature which is
approximately equal to the temperature of the gas introduced in the
preceding stage. Thus, the system shown in FIG. 1 closely
approaches isothermal operation and the high efficiencies that are
realized by such isothermal turbo-machines.
[0052] As previously mentioned, the turbo-machines of the invention
may be either axial flow or radial flow compressors or gas
turbines. An axial flow compressor in accordance with the invention
is shown in FIG. 2; a radial flow compressor is shown in FIG. 3; an
axial flow turbine is shown in FIG. 4; and a radial flow turbine is
shown in FIG. 5. In the axial flow turbo-machines, the gas is
introduced via an intake 30 as seen in the compressor in FIG. 2 and
the turbine in FIG. 4. The gas will then flow past the series of
stators S1-S5 and rotors R1-R5 in each of the stages, five of which
are shown in FIGS. 2 and 4. The direction of the gas flow is shown
by the arrows in FIGS. 2 and 4. The rotors R1-R5 are mechanically
mounted to a drive shaft 32 in any suitable known manner, such as
by way of a frustoconical element 34 as seen in FIGS. 2 and 4, so
that the rotor blades rotate with and at the same speed as the
drive shaft 32. The stators S1-S5 are stationarily mounted to the
housing 36 of the turbo-machine. In the axial turbo-machines, the
gas passes sequentially through each stage and the rotor and stator
blades in each stage. Once the gas has finally passed completely
through all of the stages, it is discharged through an exhaust
outlet 38. In the case of the compressor shown in FIG. 2, it will
be seen that the size of each stage as the gas progressively moves
through the turbo-machine decreases due to the compression of the
gas. Thus, stage 1 is larger than stage 2 which is larger than
stage 3, etc. Conversely, the size of each stage in the turbine as
shown in FIG. 4 increases as the gas progressively moves through
the turbo-machine due to the expansion of the gas in the
turbine.
[0053] The radial flow compressor shown in FIG. 3 and gas turbine
shown in FIG. 5 have similar elements in common to the axial flow
machines previously described including an intake 30, a drive shaft
32, frustoconical drive element 34, a housing 36 and an exhaust
outlet 38. The direction of the gas flow again is shown by the
arrows in FIGS. 3 and 5. The radial flow turbo-machines do differ
from the axial flow machines in certain details. One is that the
gas as it passes through the radial machines is discharged from the
rotors in a radial centrifugal manner, rather than in the axial
flow direction in the axial machines. Also, because of the radial
flow of the gas, the stators are not needed in the radial machines
and are eliminated. However, like the axial flow turbo-machines,
the sizes of the successive stages decrease toward the exhaust
outlet in the radial flow compressor shown in FIG. 3, and increase
toward the exhaust outlet in the radial flow gas turbine shown in
FIG. 5.
[0054] The axial and radial turbo-machines thus far described and
without further modifications will operate in a typical adiabatic
fashion. In adiabatic operation, as the gas moves through the
turbo-machine, its temperature and pressure will change by a
certain amount in each stage. For example, when air is introduced
to the intake 30 of the compressor it may be at ambient temperature
and pressure. However, as it progressively moves through each of
the stages 1-5 of the compressor, it will be progressively
compressed, and its temperature will elevate from stage to stage
until it is ultimately discharged from the compressor at a final
temperature and pressure which is substantially greater than the
ambient input air. Conversely, as a flow of heated gas under
pressure is introduced to a gas turbine in adiabatic operation, the
gas will expand as it moves through each of the stages, and its
pressure and temperature will drop from stage to stage so that the
gas which is ultimately discharged from the exhaust outlet 38 in an
adiabatic gas turbine will have a much greater volume and lower
pressure and temperature than the gas which was introduced
initially to the intake 30. Thus, in such purely adiabatic
turbo-machines, the temperature of the gas as it is introduced to
each successive stage is essentially the temperature of the gas as
it was discharged from the preceding stage.
[0055] The present invention is not only directed to such purely
adiabatic turbo-machines, but also to isothermal turbo-machines and
adiabatic/isothermal hybrids of such machines. An
adiabatic/isothermal hybrid is schematically depicted in FIG. 1 in
which the operation in each of the stages 1-5 is adiabatic, i.e.
increases in pressure and temperature between the intake of a given
stage and the discharge from the stage in the case of a compressor,
and decreases in pressure and temperature between the intake of a
given stage and the discharge from that stage in the case of a gas
turbine. However, in the system shown in FIG. 1, this change in
temperature is adjusted and compensated by the intercoolers C1-C5
between the stages in the case of the compressor 16 and the
interburners or combustors B1-B5 between the stages in the turbine
20. These intercoolers and interburners are preferably controlled
to cool the gas which is discharged from a stage of the compressor
16 or heat the gas which is discharged from a stage of the turbine
20 to return the gas to its initial temperature as it entered the
intake of the preceding stage and before it is introduced to the
next stage. This, in effect, adjusts the adiabatic turbo-machine
operation to perform in a manner quite similar to an isothermal
machine in which adjustments are made to the gas temperature within
the confines of each stage to maintain a constant temperature
throughout the machine as will be next discussed. This intercooler
and interburner temperature adjustment greatly improves the
efficiency of the turbo-machines. The turbo-machines in the system
of FIG. 1 may be either axial or radial flow machines.
[0056] The axial compressor shown in FIG. 2 has been modified to
function as closely as possible to a pure isothermal compressor. As
such, each of the stator blades S1-S5 includes an inlet 40 and
discharge 42 for the flow of coolant through the stator blades of
each stage. The flow of coolant is preferably controlled by a
control valve 44 which is operated by a suitable temperature sensor
46 in each stage to sense the temperature of the gas as it is
leaving the stage and adjust the flow of coolant accordingly. The
coolant may be any suitable source of coolant, including the sea
water coolant from below the thermocline as discussed with respect
to the system shown in FIG. 1.
[0057] Thus, it will be seen that in operation as the air to be
compressed in the compressor of FIG. 2 enters through the intake
30, it will be directed to the rotor blades R1 in stage 1. The
rotor blades R1 are driven by the drive shaft 32 and frustoconical
element 34 and will compress the gas to raise its pressure and
temperature in stage 1. This gas will then be directed by the
stator blades S1 at the discharge of stage 1 to the rotor blades R2
of the next stage 2. However, the coolant in the stator blades S1
in stage 1 will cool the gas which has been compressed in stage 1,
preferably to approximately the temperature that that same gas
entered stage 1 prior to compression. Thus, the gas discharged from
stage 1 will enter stage 2 ideally at the same temperature that it
entered stage 1. This is classic isothermal operation. Cooling of
the gas also occurs in the radial compressor depicted in FIG. 3. As
in the axial compressor of FIG. 2, the radial compressor shown in
FIG. 3 also includes an inlet 40 and discharge 42 for coolant, but
instead the coolant flows through a jacket 48 in the housing 36 in
each of the stages 1-5 in contrast to the circulation through the
stator blades in the axial flow compressor. Again, the coolant flow
may be controlled as desired by way of a temperature sensor 46 and
control valve 44 as previously described. In operation, the gas to
be compressed in the radial flow turbo-machine of FIG. 3 enters
through the intake 30 and is compressed by the rotor R1 in stage 1.
However, before the gas leaves stage 1, it will be cooled ideally
to the temperature that it entered stage 1 by the coolant jacket 48
in stage 1 with the flow of coolant being controlled by control
valve 44 which, in turn, is controlled by the temperature sensor
46. Thus, the gas entering stage 2 will be at the same temperature
which it entered stage 1. This is classic isothermal operation.
[0058] The axial gas turbine shown in FIG. 4 also has been modified
to function as closely as possible to a pure isothermal turbine. As
such, a fuel injector 50 is positioned between each of the stator
blades S1-S5 and their respective rotor blades R1-R5 to inject fuel
to be burned in each stage. The flow of fuel is preferably
controlled by a control valve 52 which is operated by a suitable
temperature sensor 54 in each stage to sense the temperature of the
gas as it is leaving the stage and adjust the flow of fuel
accordingly.
[0059] Thus it will be seen that in operation as the gas to propel
the turbine of FIG. 4 enters through the intake 30, it will be
directed to the stator blades S1 in stage 1. The rotor blades R1
drive the drive shaft 32 and frustoconical element 34 when they are
rotated by the gas, and the gas pressure and temperature will drop
in stage 1 as the gas passes the rotor blades R1. The gas leaving
the rotor blades R1 will then flow to the stator blades S2 and the
rotor blades R2 of the next stage 2. However, the fuel injected by
the fuel injector 50 in stage 1 will heat the gas in stage 1,
preferably to approximately the temperature that the same gas
entered stage 1 prior to expansion. Thus, the gas discharged from
stage 1 will enter stage 2 ideally at the same temperature that it
entered stage 1. Again, this is classic isothermal operation.
[0060] Heating of the gas also occurs in the radial gas turbine
depicted in FIG. 5. As in the axial turbine of FIG. 4, the radial
turbine shown in FIG. 5 also includes a fuel injector 50 for fuel
in each stage. Again, the fuel flow may be controlled as desired by
way of a control valve 52 and temperature sensor 54 and as
previously described. In operation, the gas to drive the radial
flow turbo-machine of FIG. 5 enters through the intake 30 and
drives the rotor R1 in stage 1. However, before the gas leaves
stage 1, it will be reheated ideally to the temperature that it
entered stage 1 by the fuel injector 50 in stage 1 with the flow of
fuel being controlled by control valve 52 which, in turn, is
controlled by the temperature sensor 54. Thus, the gas entering
stage 2 will be at the same temperature which it entered stage 1.
Again, this is classic isothermal operation.
[0061] It will be appreciated that although the stator blades or
the cooling jackets have been described to achieve cooling in the
case of the compressors and fuel injectors have been described to
achieve heating in the case of the gas turbines, other thermal
management elements may be employed as long as they are capable of
achieving their intended purposes and do not physically impair the
operation of the turbo-machines.
[0062] An important feature of the present invention is the manner
in which the respective stages are sized relative to each
other.
[0063] In the present invention one stage is selected to be a
master or model stage and it is designed to be of optimal
efficiency based upon given operating parameters for the
turbo-machine. These operating parameters include shaft speed,
pressure ratio, type of gas processing (adiabatic, isothermal,
etc.), temperature ratio and gas density ratio. Once the optimum
design features and geometric shape of the master stage is
determined, this master stage then serves as a model for the design
features and geometric shape of the remaining stages. The remaining
stages will then have substantially the same design features and
geometric shape as the master stage, but will be sized differently
relative to the parameters in accordance with the formula of the
invention.
[0064] In accordance with the invention, the ratio of linear
dimensions L of two adjacent stages is determined by the formula 5
L T = D T 3
[0065] where D is the gas density ratio of the master stage (and
each of the successive remaining stages). The gas density ratio D
is governed by the formula 6 D = ( P I / P O T I / T O )
[0066] where P.sub.I is the absolute pressure of the gas entering
the stage in question, P.sub.O is the absolute pressure of the gas
as discharged from the stage in question, T.sub.I is the absolute
temperature of the gas entering the stage in question, and T.sub.O
is the absolute temperature of the gas as discharged from the stage
in question.
[0067] Once the ratio of linear dimensions L is determined for a
given turbo-machine, each and every linear dimension of a given
stage is multiplied by this ratio to determine the comparable
linear dimension in the next successive stage. For example and with
reference to FIG. 6, if the blade length B1 in a given stage of a
rotor and/or stator would be 10 inches, and the blade width Rw of
the rotor and/or stator is one inch, and the linear dimension ratio
L is 1.201, the comparable blade length B1 in the next successive
stage will be 10.times.1.201=12.01 inches, the comparable blade
width Bw will be 1.times.1.201=1.201 inches, and the geometric
shape of the respective blades of each stage will be identical to
each other. This is also true of all of the other linear dimensions
in each of the stages, such as for example with reference to FIG.
2, the distance between the rotors and stators a, the length of the
stage b, the radius of the frustoconical element c, the distance
between the frustoconical element and the housing d, etc.
[0068] The master stage as discussed herein may be anywhere in the
multistage turbo-machine. It may be the first stage in which case
all subsequent stages increase or decrease in size of corresponding
linear dimensions depending on whether the turbo-machine is a
compressor or gas turbine. The master stage also may be one of the
middle stages, in which case the stages on opposite sides of the
master stage both decrease and increase accordingly in linear
dimensions.
EXAMPLE 1
[0069] The following calculations are presented by way of example
for the sizing of a gas turbine having five stages, and operating
conditions including an initial intake gas pressure of 32 atm
absolute, a pressure ratio (P.sub.I/P.sub.O) of 2, and an initial
intake gas temperature of 1500.degree. R which would decrease to
1300.degree. R in the first stage if the temperature was not
adjusted. The linear dimension ratio L of such turbine under such
operating conditions as calculated in accordance with the invention
are set forth in the following Tables 1A-1C together with the
manner in which the linear dimension ratio L has been calculated
for three types of gas turbine: (1) pure isothermal, (2) hybrid
(adiabatic/isothermal), and (3) pure adiabatic. In addition and by
way of example, the actual lengths of one of the components of each
stage are calculated for each stage and set forth in the Tables
1A-1C, e.g. for a blade length B1 as shown in FIG. 6 of 10.00
inches for the master stage.
1TABLE 1A Gas Turbine - Pure Isothermal Stage P.sub.Iatm abs
P.sub.0atm abs T.sub.I.degree.R T.sub.0.degree.R Calculations 7 P I
/ P O T I / T O 3 Ratio of Linear Dimensions L Blade Length Bl 1 32
16 1500 1500 8 32 / 16 1500 / 1500 3 = 2 1 3 = 1.25999 10.00 2 16 8
1500 1500 9 16 / 8 1500 / 1500 3 = 2 1 3 = 1.25999 12.60 3 8 4 1500
1500 10 8 / 4 1500 / 1500 3 = 2 1 3 = 1.25999 15.88 4 4 2 1500 1500
11 4 / 2 1500 / 1500 3 = 2 1 3 = 1.25999 20.00 5 2 1 1500 1500 12 2
/ 1 1500 / 1500 3 = 2 1 3 = 1.25999 25.20
[0070]
2TABLE 1B Gas Turbine - Hybrid (Adiabatic/Isothermal) Stage
P.sub.Iatm abs P.sub.0atm abs T.sub.I.degree.R T.sub.0.degree.R
Calculations 13 P I / P O T I / T O 3 Ratio of Linear Dimensions L
Blade Length Bl 1 32 16 1500 1300 14 32 / 16 1500 / 1300 3 = 2
1.1538 3 = 1.7334 3 = 1.201 10.00 2 16 8 1500 1300 15 16 / 8 1500 /
1300 3 = 2 1.1538 3 = 1.7334 3 = 1.201 12.01 3 8 4 1500 1300 16 8 /
4 1500 / 1300 3 = 2 1.1538 3 = 1.7334 3 = 1.201 14.42 4 4 2 1500
1300 17 4 / 2 1500 / 1300 3 = 2 1.1538 3 = 1.7334 3 = 1.201 17.32 5
2 1 1500 1300 18 2 / 1 1500 / 1300 3 = 2 1.1538 3 = 1.7334 3 =
1.201 20.81
[0071]
3TABLE 1C Gas Turbine - Pure Adiabatic Stage P.sub.Iatm abs
P.sub.0atm abs T.sub.I.degree.R T.sub.0.degree.R Calculations 19 P
I / P O T I / T O 3 Ratio of Linear Dimensions L Blade Length Bl 1
32 16 1500 1300 20 32 / 16 1500 / 1300 3 = 2 1.1538 3 = 1.7334 3 =
1.201 10.00 2 16 8 1300 1127 21 16 / 8 1300 / 1127 3 = 2 1.1538 3 =
1.7334 3 = 1.201 12.01 3 8 4 1127 977 22 8 / 4 1127 / 977 3 = 2
1.1538 3 = 1.7334 3 = 1.201 14.42 4 4 2 977 846 23 4 / 2 977 / 846
3 = 2 1.1538 3 = 1.7334 3 = 1.201 17.32 5 2 1 847 734 24 2 / 1 847
/ 734 3 = 2 1.1538 3 = 1.7334 3 = 1.201 20.81
EXAMPLE 2
[0072] The following calculations are presented by way of example
for the sizing of a gas compressor, similar to the gas turbine of
Example 1, having five stages, and operating conditions including
an initial intake gas pressure of 1 atm absolute, a pressure ratio
(P.sub.I/P.sub.O) of 0.5, and an initial intake gas temperature of
530.degree. R which would increase to 612.degree. R in the first
stage if the temperature was not adjusted. The linear dimension
ratio L of such compressor under such operating conditions as
calculated in accordance with the invention are set forth in the
following Tables 2A-2C together with the manner in which the linear
dimension ratio L has been calculated for three types of gas
compressor: (1) pure isothermal, (2) hybrid (adiabatic/isothermal),
and (3) pure adiabatic. In addition and by way of example, the
actual lengths of one of the components of each stage are
calculated for each stage and set forth in Tables 2A-2C, e.g. for a
blade length B1 as shown in FIG. 6 of 10.00 inches for the master
stage.
4TABLE 2A Gas Compressor - Pure Isothermal Stage P.sub.Iatm abs
P.sub.0atm abs T.sub.I.degree.R T.sub.0.degree.R Calculations 25 P
I / P O T I / T O 3 Ratio of Linear Dimensions L Blade Length Bl 1
1 2 530 530 26 1 / 2 530 / 530 3 = .5 1 3 = 0.7937 10.00 2 2 4 530
530 27 2 / 4 530 / 530 3 = .5 1 3 = 0.7937 7.94 3 4 8 530 530 28 4
/ 8 530 / 530 3 = .5 1 3 = 0.7937 6.30 4 8 16 530 530 29 8 / 16 530
/ 530 3 = .5 1 3 = 0.7937 5.00 5 16 32 530 530 30 16 / 32 530 / 530
3 = .5 1 3 = 0.7937 3.97
[0073]
5TABLE 2B Gas Compressor - Hybrid (Adiabatic/Isothermal) Stage
P.sub.Iatm abs P.sub.0atm abs T.sub.I.degree.R T.sub.0.degree.R
Calculations 31 P I / P O T I / T O 3 Ratio of Linear Dimensions L
Blade Length Bl 1 1 2 530 612 32 1 / 2 530 / 612 3 = .5 0.866 3 =
0.577 3 = 0.833 10.00 2 2 4 530 612 33 2 / 4 530 / 612 3 = .5 0.866
3 = 0.577 3 = 0.833 8.33 3 4 8 530 612 34 4 / 8 530 / 612 3 = .5
0.866 3 = 0.577 3 = 0.833 6.94 4 8 16 530 612 35 8 / 16 530 / 612 3
= .5 0.866 3 = 0.577 3 = 0.833 5.78 5 16 32 530 612 36 16 / 32 530
/ 612 3 = .5 0.866 3 = 0.577 3 = 0.833 4.81
[0074]
6TABLE 2C Gas Compressor - Pure Adiabatic Stage P.sub.Iatm abs
P.sub.0atm abs T.sub.I.degree.R T.sub.0.degree.R Calculations 37 P
I / P O T I / T O 3 Ratio of Linear Dimensions L Blade Length Bl 1
1 2 530 612 38 1 / 2 530 / 612 3 = .5 0.866 3 = 0.577 3 = 0.833
10.00 2 2 4 612 707 39 2 / 4 612 / 707 3 = .5 0.866 3 = 0.577 3 =
0.833 8.33 3 4 8 707 816 40 4 / 8 707 / 816 3 = .5 0.866 3 = 0.577
3 = 0.833 6.94 4 8 16 816 942 41 8 / 16 816 / 942 3 = .5 0.866 3 =
0.577 3 = 0.833 5.78 5 16 32 942 1088 42 16 / 32 942 / 1088 3 = .5
0.866 3 = 0.577 3 = 0.833 4.81
[0075] The foregoing description, with respect to stage sizing is
effective where additional tributary gasses or the like are not
added to or removed from the gas stream in one or more of the
stages or in the intercoolers C1-C5 or interburners B1-B5 between
the stages, i.e. the only gasses which leave the turbo-machine at
its discharge are the same gasses which entered the turbo-machine
at its inlet. Such turbo-machines would include turbo-machines,
such as the compressors in which the gasses flow through
intercoolers C.sub.1-C.sub.5 as shown in FIG. 1 without the
addition of tributary gasses in the intercoolers, compressors in
which the coolant flows only through the turbo-machine blades or
through jackets as shown in FIGS. 2 and 3, and/or turbines having
interburners B.sub.1-B.sub.5 as shown in FIG. 1 without the
addition of tributary gasses in the interburners. In each of these
turbo-machines, gasses are not added to the gas which was initially
introduced to the inlet of the compressors or turbines.
[0076] However, in some instances it may be desired to add
tributary gasses to one or more of the stages, the intercoolers
and/or interburners during the flow of the primary gas stream
through the turbo-machine, or to remove some of the primary gas
stream from individual stages, the intercoolers and/or
interburners. This may occur for example in an isothermal gas
turbine, such as depicted in FIGS. 4 and 5, in which fuel is
injected into and combusted in each stage to maintain the
temperature substantially constant from the inlet to the discharge
of each stage. This may also occur where fluids are employed to
cool a turbine blade itself to prevent the temperature of the blade
from exceeding a desired limit. Such cooling may be accomplished by
the tributary fuel itself which is injected into the stage, such as
with endothermic fuels, evaporative cooling of the blade surfaces
or using cryogenic fuels such as liquified hydrogen. This may also
occur in compressors where it is desired to operate under
isothermal conditions or a tributary stream is removed from one or
more intermediate stages at a lower pressure before the final
discharge from the compressor. In each of these cases where the
tributary gas is added or removed from one or more stages, the
ratio of linear dimensions as previously described should
preferably be adjusted to compensate for the tributary fluids which
have been added or removed. When the ratio of linear dimensions has
been computed in accordance with the present invention to take into
account the tributary additions or withdrawals to or from the
turbo-machine stages or between the stages, each stage will develop
the same amount of shaft energy per unit weight of gas flow and
each stage will have the same efficiency which should be optimum it
the master stage efficiency has been designed to be optimum.
[0077] By way of example a schematic depiction of a gas
turbo-machine having tributary additions to each stage appears in
FIG. 7. The schematic gas turbo-machine is shown as having five
stages, although it will be appreciated that it may have more or
less than this number of stages. As shown in FIG. 7, the gas
introduced to the intake to stage 1 has a molar volume flow rate in
moles/sec of M.sub.A. The intake to stage 1 is typically the
exhaust of a combustor in which a fuel, such as a hydrocarbon or
hydrogen, is burned with air or oxygen, and this exhaust is
introduced to the intake of stage 1 at an elevated temperature and
pressure. Each mole of the hot gasses which is introduced to the
intake into stage 1, will have a given volume at a given
temperature and pressure. For example, if the fuel is hydrogen
which has been oxidized with oxygen and it is assumed that the
combustion is 100% efficient, the content of the gas combustion
products introduced to the intake of stage 1 will be H.sub.2O.
There may also be an excess of O.sub.2 entering the intake of stage
1 to provide for the oxidation of the tributary additions of fuel
which when oxidized produces a net molar volume M.sub.B of the
tributary fluids in each stage in moles/sec. Each gram-molecular
volume of a gas is known to be 22.4 liters at STP. Accordingly,
each gram-mole of H.sub.2O and/or of O.sub.2 introduced to the
intake of stage 1, or each gram-mole of H.sub.2O which is produced
where the tributary fuel burned in a stage is H.sub.2 will have a
volume of 22.4 liters S.T.P. adjusted to the temperature and
pressure of the gas.
[0078] As shown in FIG. 7, a tributary gas is introduced to each
stage which will generate a molar volume flow rate of M.sub.B in
moles/sec in each stage. Where the fuel which has been combusted to
form the gas which is introduced to the intake of stage 1 is
hydrogen, the tributary gas would typically also be hydrogen. In an
isothermal gas turbine in which H.sub.2 would be burned to maintain
a substantially constant temperature is desired of the gas
introduced to a stage and the gas discharged from a stage, the
input of H.sub.2 would be adjusted to an amount that when oxidized
or burned by the excess O.sub.2 Will generate sufficient heat to
maintain the substantially constant temperature across the stage.
Thus, the flow rate of the gas discharged from stage 1 and
introduced to the intake of stage 2 would be M.sub.A+M.sub.B
mole/sec.
[0079] Again, a molar volume amount of a tributary input M.sub.B is
generated within stage 2 in order to maintain the temperature
constant across stage 2 despite the fact that the gasses are
expanding in the stage. Thus, the gas discharged from stage 2 and
introduced to the intake of stage 3 will have a molar volume flow
rate of (M.sub.A+2M.sub.B) moles/sec, which is the total tributary
gas input for stages 1 and 2 and the molar intake of stage 1
(M.sub.A).
[0080] Likewise, the gas discharged from stage 3 and introduced to
stage 4 will have a molar volume of (M.sub.A+3M.sub.B); the gas
discharged from stage 4 and introduced to the intake of stage 5
will have a molar volume of (M.sub.A+4M.sub.B), and finally, the
gas discharged from stage 5 and from the gas turbo-machine where
the turbo-machine only has five stages will have a molar volume of
(M.sub.A+5M.sub.B). The foregoing also applies where the tributary
gasses are added between the stages to the intercoolers C1-C5
and/or the interburners B1-B5.
[0081] As previously mentioned, an important feature of the present
invention is the manner in which the respective stages are sized
relative to each other in such gas turbo-machines as last described
in which a tributary flow rate of gas M.sub.B is generated in or
removed from or between one or more of the turbo-machine stages. In
such tributary gas turbo-machines, the molar volume flow-rate
M.sub.B of the additional tributary matter must be and is taken
into account in the design and sizing of the stages for optimal
efficiency in the present invention.
[0082] In the tributary gas turbo-machines, one stage is again
selected to be a master or model stage and it is designed to be of
optimal efficiency based upon given operating parameters for the
turbo-machine, such as shaft speed, pressure ratio, type of gas
processing (adiabatic, isothermal, etc.), and temperature ratio.
Once the optimum design features and geometric shape of the master
stage are determined, this master stage then serves as a model for
the design features and geometric shape of the remaining stages.
The remaining stages will then have substantially the same design
features and geometric shape as the master stage, but will be sized
differently relative to the parameters in accordance with the
formula of the invention.
[0083] In accordance with the invention, the ratio of linear
dimensions L.sub.T of a stage which is downstream of the master
stage where tributary gasses are added or removed to or from one or
more stages and/or between stages is determined by the formula 43 L
T = D T 3
[0084] where:
[0085] D.sub.T is the gas density ratio of each stage. The gas
density ratio D.sub.T of each stage is governed by the formula
where: 44 D T = ( M A + M B tot M A ) ( P I / P O T I / T O ) n -
1
[0086] M.sub.A=molar volume flow rate to intake of stage 1,
moles/sec;
[0087] M.sub.B=tributary volume flow rate generated in or between
each stage, moles/sec;
[0088] M.sub.B tot=total tributary volume added to or between all
preceding stages, moles/sec;
[0089] P.sub.I=absolute pressure of gas entering stage in
question;
[0090] P.sub.O=absolute pressure of gas leaving stage in
question;
[0091] T.sub.I=absolute temperature of gas entering stage in
question;
[0092] T.sub.O=absolute temperature of gas leaving stage in
question; and
[0093] n=number of the stage in question.
[0094] Once the linear dimensions and spacing of the components are
determined in an optimal master stage, the ratio of linear
dimensions L.sub.T is determined for a given downstream stage of a
turbo-machine having tributary flow to or from one or more of the
stages or between the stages. The linear dimensions of the
components and their spacing in the master stage are then
multiplied by the ratio of linear dimensions L.sub.T for that stage
to determine the comparable linear dimensions of the components and
spacing for each subsequent stage. For example and with reference
to FIG. 6, if the blade length B1 in the master stage of a rotor
and/or stator is 10 inches, the blade width Rw of the rotor and/or
stator is 1 inch, and the linear dimension ratio L.sub.T for the
next stage is 1.264, the comparable blade length B1 in the next
stage will be 10.times.1.264 inches, the comparable blade width Bw
will be 1.times.1.264 inches, and the geometric shape of the
respective blades of each stage will be identical to each other.
This is also true of all of the other linear dimensions in the
second stage, such as for example with reference to FIG. 2, the
distance a between the rotor R1 and stator S1, the length of the
stage b, the radius c of the frustoconical element 34, the distance
d between the frustoconical element 34 and the housing 36, etc.
[0095] The master stage as discussed herein again may be anywhere
in the multistage turbo-machine as previously mentioned. It may be
the first stage in which case all subsequent stages increase in
size in a turbine and vice versa in the case of a compressor. The
master stage also may be one of the middle stages, in which case
the stages downstream will increase and upstream will decrease
accordingly in linear dimensions in the case of a turbine and vice
versa in the case of a compressor.
EXAMPLE 3
[0096] The following calculations are presented by way of example
for the sizing of a tributary gas turbine having five stages, and
operating conditions including an initial intake gas pressure of 32
atm absolute, a pressure ratio (P.sub.I/P.sub.O) of 2, an initial
intake gas molar volume flow rate M.sub.A of 100 units of moles/sec
and a tributary gas molar volume flow rate of M.sub.B of 1 unit of
moles/sec generated in each stage. The linear dimension ratios
L.sub.T of each of the five turbine stages under such operating
conditions as calculated in accordance with the invention are set
forth in the following Tables 3A-3C together with the manner in
which the linear dimension ratio L.sub.T has been calculated for
each of the stages. In addition and by way of example, the actual
lengths of one of the components of each stage are calculated for
each stage and set forth in Tables 3A-3C, i.e., for a blade length
B1 as shown in FIG. 6 of 10.00 inches for a first master stage.
7TABLE 3A Gas Turbine with Tributary Flow - Pure Isothermal Stage n
P.sub.Iatm abs P.sub.0atm abs T.sub.I.degree.R T.sub.0.degree.R
Calculations 45 ( M A + M B tot M A ) ( P I / P O T I / T O ) n - 1
3 Ratio of Linear Dimensions L.sub.T Blade Length Bl 1 32 16 1500
1500 46 ( 100 + 0 100 ) ( 32 / 16 1500 / 1500 ) 1 - 1 3 = 1 .times.
2 0 3 = 1 .times. 1 3 = 1 3 = 1.0000 10.00 2 16 8 1500 1500 47 (
100 + 1 100 ) ( 16 / 8 1500 / 1500 ) 2 - 1 3 = 1.01 .times. 2 1 3 =
1.01 .times. 2 3 = 2.02 3 = 1.264 12.64 3 8 4 1500 1500 48 ( 100 +
2 100 ) ( 8 / 4 1500 / 1500 ) 3 - 1 3 = 1.02 .times. 2 2 3 = 1.02
.times. 4 3 = 4.08 3 = 1.5978 15.98 4 4 2 1500 1500 49 ( 100 + 3
100 ) ( 4 / 2 1500 / 1500 ) 4 - 1 3 = 1.03 .times. 2 3 3 = 1.03
.times. 8 3 = 8.24 3 = 2.0197 20.20 5 2 1 1500 1500 50 ( 100 + 4
100 ) ( 2 / 1 1500 / 1500 ) 5 - 1 3 = 1.04 .times. 2 4 3 = 1.04
.times. 16 3 = 16.14 3 = 2.5528 25.53
[0097]
8TABLE 3B Gas Turbine with Tributary Flow - Hybrid
(Adiabatic/Isothermal) Stage n P.sub.Iatm abs P.sub.0atm abs
T.sub.I.degree.R T.sub.0.degree.R Calculations 51 ( M A + M B tot M
A ) ( P I / P O T I / T O ) n - 1 3 Ratio of Linear Dimen- sions
L.sub.T Blade Length Bl 1 32 16 1500 1300 1.00 10.0 52 ( 100 + 0
100 ) ( 32 / 16 1500 / 1300 ) 1 - 1 3 = 1 .times. ( 2 1.15 ) 0 3 =
1 .times. 1.74 0 3 = 1 .times. 1 3 = 1 3 = 2 16 8 1500 1300 1.21
12.1 53 ( 100 + 1 100 ) ( 16 / 18 1500 / 1300 ) 2 - 1 3 = 1.01
.times. ( 2 1.15 ) 1 3 = 1.01 .times. 1.74 1 3 = 1.01 .times. 1.74
3 = 1.76 3 = 3 8 4 1500 1300 1.46 14.6 54 ( 100 + 2 100 ) ( 8 / 4
1500 / 1300 ) 3 - 1 3 = 1.02 .times. ( 2 1.15 ) 2 3 = 1.02 .times.
1.74 2 3 = 1.02 .times. 3.03 3 = 3.09 3 = 4 4 2 1500 1300 1.76 17.6
55 ( 100 + 3 100 ) ( 4 / 2 1500 / 1300 ) 4 - 1 3 = 1.03 .times. ( 2
1.15 ) 0 3 = 1.03 .times. 1.74 3 3 = 1.03 .times. 5.27 3 = 5.43 3 =
5 2 1 1500 1300 2.12 21.2 56 ( 100 + 4 100 ) ( 2 / 1 1500 / 1300 )
5 - 1 3 = 1.04 .times. ( 2 1.15 ) 4 3 = 1.04 .times. 1.74 4 3 =
1.04 .times. 9.17 3 = 9.54 3 =
[0098]
9TABLE 3C Gas Turbine with Tributary Flow - Pure Adiabatic Stage n
P.sub.Iatm abs P.sub.0atm abs T.sub.I.degree.R T.sub.0.degree.R
Calculations 57 ( M A + M B tot M A ) ( P I / P O T I / T O ) n - 1
3 Ratio of Linear Dimen- sions L.sub.T Blade Length Bl 1 32 16 1500
1300 58 ( 100 + 0 100 ) ( 32 / 16 1500 / 1300 ) 1 - 1 3 = 1 .times.
( 2 1.15 ) 0 3 = 1 .times. 1.74 0 3 = 1 .times. 1 3 = 1 3 = 1.00
10.0 2 16 8 1300 1127 59 ( 100 + 1 100 ) ( 16 / 8 1300 / 1127 ) 2 -
1 3 = 1.01 .times. ( 2 1.15 ) 1 3 = 1.01 .times. 1.74 1 3 = 1.01
.times. 1.74 3 = 1.76 3 = 1.21 12.1 3 8 4 1127 977 60 ( 100 + 2 100
) ( 8 / 4 1127 / 977 ) 3 - 1 3 = 1.02 .times. ( 2 1.15 ) 2 3 = 1.02
.times. 1.74 2 3 = 1.02 .times. 3.03 3 = 3.09 3 = 1.46 14.6 4 4 2
977 846 61 ( 100 + 3 100 ) ( 4 / 2 977 / 846 ) 4 - 1 3 = 1.03
.times. ( 2 1.15 ) 3 3 = 1.03 .times. 1.74 3 3 = 1.03 .times. 5.27
3 = 5.43 3 = 1.76 17.6 5 2 1 846 734 62 ( 100 + 4 100 ) ( 2 / 1 846
/ 734 ) 5 - 1 3 = 1.04 .times. ( 2 1.15 ) 4 3 = 1.04 .times. 1.74 1
3 = 1.04 .times. 9.17 3 = 9.54 3 = 2.12 21.2
EXAMPLE 4
[0099] The following calculations are presented by way of example
for the sizing of a tributary gas compressor having five stages,
and operating conditions including an initial intake gas pressure
of 1 atm absolute, a pressure ratio (P.sub.I/P.sub.O) of 0.5, an
initial intake gas molar volume flow rate M.sub.A of 100 units of
moles/sec and a tributary gas output of a molar volume flow rate
M.sub.B of 5 units of moles/sec from the third stage, but no
tributary input or output in the remaining stages. The linear
dimension ratios L.sub.T of each of the five compressor stages
under such operating conditions as calculated in accordance with
the invention are set forth in the following Tables 4A-4C together
with the manner in which the linear dimension ratio L.sub.T has
been calculated for each of the stages. In addition and by way of
example, the actual lengths of one of the components of each stage
are calculated for each stage and set forth in Tables 4A-4C, i.e.,
for a blade length B1 as shown in FIG. 6 of 10.00 inches for a
first master stage.
10TABLE 4A Gas Compressor with Tributary Outflow From Third Stage -
Pure Isothermal Stage n P.sub.Iatm abs P.sub.0atm abs
T.sub.I.degree.R T.sub.0.degree.R Calculations 63 ( M A + M B tot M
A ) ( P I / P O T I / T O ) n - 1 3 Ratio of Linear Dimensions
L.sub.T Blade Length Bl 1 1 2 530 530 64 ( 100 0 100 ) ( 1 / 2 530
/ 530 ) 1 - 1 3 = 1 .times. ( .5 1 ) 0 3 = 1 .times. 1 3 = 1 3 =
1.00 10.0 2 2 4 530 530 65 ( 100 0 100 ) ( 2 / 4 530 / 530 ) 2 - 1
3 = 1 .times. ( .5 1 ) 1 3 = 1 .times. 5 3 = .5 3 = 0.794 7.94 3 4
8 530 530 66 ( 100 0 100 ) ( 4 / 8 530 / 530 ) 3 - 1 3 = 1 .times.
( .5 1 ) 2 3 = 1 .times. .25 3 = .25 3 = 0.630 6.30 4 8 16 530 530
67 ( 100 - 5 100 ) ( 8 / 16 530 / 530 ) 4 - 1 3 = 95 100 .times. (
.5 1 ) 3 3 = .95 .times. 1.25 3 = 0.119 3 = 0.492 4.92 5 16 32 530
530 68 ( 100 - 5 100 ) ( 16 / 32 530 / 530 ) 5 - 1 3 = 95 100
.times. ( .5 1 ) 4 3 = .95 .times. .0625 3 = 0.059 3 = 0.389
3.89
[0100]
11TABLE 4B Gas Compressor With Tributary Outflow From Third
Stage-Hybrid (Adiabatic/Isothermal) Stage n P.sub.Iatm abs
P.sub.Oatm abs T.sub.I.sup.OR T.sub.O.sup.OR Calculations 69 ( M A
+ M B tot M A ) ( P I / P O T I / T O ) n - 1 3 Ratio of Linear
Dimensions L.sub.T Blade Length B1 1 1 2 530 612 70 ( 100 0 100 ) (
1 / 2 530 / 612 ) 1 - 1 3 = 1 .times. ( .5 .866 ) 0 3 = 1 .times.
0.577 0 3 = 1 .times. 1 3 = 1 3 = 1.00 10.00 2 2 4 530 612 71 ( 100
0 100 ) ( 2 / 4 530 / 612 ) 2 - 1 3 = 1 .times. ( .5 .866 ) 1 3 = 1
.times. 0.577 1 3 = 1 .times. 0.577 1 3 = 0.577 3 = 0.833 8.33 3 4
8 530 612 72 ( 100 0 100 ) ( 4 / 8 530 / 612 ) 2 - 1 3 = 1 .times.
( .5 .866 ) 2 3 = 1 .times. 0.577 2 3 = 1 .times. 0.333 1 3 = 0.333
3 = 0.693 6.93 4 8 16 530 612 73 ( 100 - 5 100 ) ( 8 / 16 530 / 612
) 4 - 1 3 = 95 100 .times. ( .5 .866 ) 3 3 = .95 .times. 0.577 3 3
= .95 .times. 0.192 3 = 0.182 3 = 0.567 5.67 5 16 32 530 612 74 (
100 - 5 100 ) ( 16 / 32 530 / 612 ) 5 - 1 3 = 95 100 .times. ( .5
.866 ) 4 3 = .95 .times. 0.577 4 3 = .95 .times. 0.111 3 = 0.105 3
= 0.472 4.72
[0101]
12TABLE 4C Gas Compressor With Tributary Outflow From Third
Stage-Pure Adiabatic Stage n P.sub.Iatm abs P.sub.Oatm abs
T.sub.I.sup.OR T.sub.O.sup.OR Calculations 75 ( M A + M B tot M A )
( P I / P O T I / T O ) n - 1 3 Ratio of Linear Di- men- sions
L.sub.T Blade Length B1 1 1 2 530 612 76 ( 100 0 100 ) ( 1 / 2 530
/ 612 ) 1 - 1 3 = 1 .times. ( .5 .866 ) 0 3 = .95 .times. 0.577 0 3
= 1 .times. 1 3 = 1 3 = 1.00 10.00 2 2 4 612 707 77 ( 100 0 100 ) (
2 / 4 612 / 707 ) 2 - 1 3 = 1 .times. ( .5 .866 ) 1 3 = 1 .times.
0.577 1 3 = 1 .times. 0.577 3 = 0.577 3 = 0.833 8.33 3 4 8 707 816
78 ( 100 0 100 ) ( 4 / 8 707 / 816 ) 3 - 1 3 = 1 .times. ( .5 .866
) 2 3 = 1 .times. 0.577 2 3 = 1 .times. 0.333 3 = 0.333 3 = 0.693
6.93 4 8 16 816 942 79 ( 100 - 5 100 ) ( 8 / 16 816 / 942 ) 2 - 1 3
= 95 100 .times. ( .5 .866 ) 3 3 = .95 .times. 0.577 3 3 = .95
.times. 0.192 3 = 0.182 3 = 0.567 5.67 5 16 32 942 1088 80 ( 100 -
5 100 ) ( 16 / 32 942 / 1088 ) 5 - 4 3 = 95 100 .times. ( .5 .866 )
4 3 = .95 .times. 0.577 4 3 = .95 .times. 0.111 3 = 0.105 3 = 0.472
4.72
[0102] A few observations should be noted with respect to Tables
1A-C and 2A-C and Tables 3A-C and 4A-C. 1. One observation will be
that the ratio of linear dimensions L as calculated in the Tables
1A-C and 2A-C are from one stage to the next, i.e. the change in
linear dimensions from stage 1 to stage 2, the change in linear
dimensions from stage 2 to stage 3, etc. However, the ratio of
linear dimensions L.sub.T as calculated in Tables 3A-C and 4A-C is
the change in linear dimensions of any given stage to the master
stage, i.e. the linear change from stage 1 to stage 2, the linear
change from stage 1 to stage 3, the linear change from stage 1 to
stage 4, etc. The calculations of linear change from the fixed
reference stage 1 is necessitated in Tables 3A-C and 4A-C by the
tributary inflow or outflow to or from one or more of the
downstream stages or between the downstream stages. Such tributary
inflow or outflow does not exist in the turbines or compressors of
Tables 1A-C and 2A-C and, therefore, the relative linear change
between any given stage and its next stage is a constant. That is
not the case, however, in the tributary turbo-machines where
tributary streams are added or removed at intermediate locations
over the course of flow through the turbo-machines.
[0103] It will also be observed upon consideration of the
calculations in Tables 3A-3C and 4A-4C that the formula used to
calculate the ratio of linear dimensions L.sub.T in those tables
will also work equally well in turbo-machines which have no
tributary inflow to or outflow from one or more of the respective
stages or between the stages. In that case, M.sub.B tot is zero for
each stage and 81 M A M A = 1 1
[0104] so that the remainder of the equation is simply the gas
density ratio D as set forth in the equation used to calculate the
ratio of linear dimensions in the non-tributary turbo-machines of
Tables 1A-1C and 2A-2C.
[0105] It will be understood that the preferred embodiments of the
present invention as have been described are merely illustrative of
the principles of the present invention. Numerous modifications may
be made by those skilled in the art without departing from the true
spirit and scope of the invention.
* * * * *