U.S. patent number 3,696,637 [Application Number 05/194,856] was granted by the patent office on 1972-10-10 for method and apparatus for producing refrigeration.
This patent grant is currently assigned to Air Products and Chemicals, Inc.. Invention is credited to Leif A. Ness, Edmund P. Thomas.
United States Patent |
3,696,637 |
Ness , et al. |
October 10, 1972 |
METHOD AND APPARATUS FOR PRODUCING REFRIGERATION
Abstract
Refrigeration system including a multi-stage compressor and two
work expansion engines of the turbine type with the impellers of
the expansion engines and the impeller of the final stage of the
compressor being mounted on a common shaft. The work developed in
the expansion engines provides the total power required for the
final stage of the compressor and the final stage of the compressor
provides the pressurized gas expanded in both expansion
engines.
Inventors: |
Ness; Leif A. (Macungie,
PA), Thomas; Edmund P. (Bethlehem, PA) |
Assignee: |
Air Products and Chemicals,
Inc. (Allentown, PA)
|
Family
ID: |
26890471 |
Appl.
No.: |
05/194,856 |
Filed: |
November 2, 1971 |
Related U.S. Patent Documents
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Application
Number |
Filing Date |
Patent Number |
Issue Date |
|
|
752998 |
Aug 15, 1968 |
3657898 |
|
|
|
Current U.S.
Class: |
62/402; 62/88;
62/910 |
Current CPC
Class: |
F25J
3/04278 (20130101); F25J 1/0279 (20130101); F25B
9/06 (20130101); F25J 3/04866 (20130101); F25J
3/04224 (20130101); F25J 1/0234 (20130101); F25J
1/0236 (20130101); F25J 1/0208 (20130101); F25J
1/0288 (20130101); F25J 1/004 (20130101); F25J
3/04381 (20130101); F25J 1/0037 (20130101); F25J
3/04393 (20130101); F25J 1/0015 (20130101); F25J
3/04357 (20130101); F25J 2240/02 (20130101); F25J
2270/90 (20130101); F25J 2270/06 (20130101); F25J
2230/20 (20130101); Y10S 62/91 (20130101) |
Current International
Class: |
F25J
1/00 (20060101); F25B 9/06 (20060101); F25d
009/00 () |
Field of
Search: |
;62/86,87,88,402 |
References Cited
[Referenced By]
U.S. Patent Documents
Primary Examiner: Wye; William J.
Parent Case Text
This a division of application Ser. No. 752,998, filed 8-15-68, now
U.S. Pat. No. 3,657,898.
Claims
We claim:
1. A compressor-turbine assembly comprising:
a. a rotatable shaft;
b. a compressor impeller secured to said shaft;
c. casing means forming a compression chamber surrounding said
compressor impeller and forming inlet and discharge passageways
communicating with said compression chamber;
d. a first turbine impeller secured to said shaft axially spaced on
one side of said compressor impeller;
e. casing means forming an expansion chamber surrounding said first
turbine impeller and forming inlet and discharge passageways
communicating with said first expansion chamber;
f. a second turbine impeller secured to said shaft axially spaced
on the other side of said compressor impeller;
g. casing means forming an expansion chamber surrounding said
second turbine impeller and forming inlet and discharge passageways
communicating with said second expansion chamber;
h. a first set of adjustable vanes located in the inlet passageway
communicating with said first expansion chamber, and a second set
of adjustable vanes located in the inlet passageway communicating
with said second expansion chamber; and
i. radially and circumferentially extending spaces forming annular
chambers between the periphery of each of said turbine impellers
and the adjustable vanes associated therewith.
2. The compressor-turbine assembly as claimed in claim 1 wherein
said inlet passageways communicating with said first and second
turbine expansion chambers comprise radially extending passages
communicating with the periphery of said turbine impellers.
3. The compressor-turbine assembly as claimed in claim 1 wherein
said discharge passageways communicating with said first and second
turbine expansion chambers comprise discharge passages extending
co-axially with said shaft and directed to discharge the expanded
gas in opposite directions away from said compressor casing
means.
4. The compressor-turbine assembly as claimed in claim 1 wherein
said compressor inlet passageway comprises a radially inwardly and
axially extending passage, said compressor outlet passageway
comprises a radially outwardly extending passageway, said first and
second turbine inlet passageways comprise radially inwardly
extending passages communicating with said vanes, and said
discharge passageways comprise axially extending passages directed
in opposite directions away from said compressor.
5. The compressor-turbine assembly as claimed in claim 1 wherein
said first and second turbine impellers are of different size
having different tip velocities, said first and second adjustable
vanes and said annular chambers comprising means for matching the
tangential angles of the inlet gases to the different tip speeds of
the respective turbine impellers.
6. The compressor-turbine assembly as claimed in claim 5 including
means for simultaneously actuating the first and second sets of
adjustable vanes.
7. The compressor-turbine assembly as claimed in claim 5 including
means for independently adjusting said first and second sets of
adjustable vanes.
8. The compressor-turbine assembly as claimed in claim 5 wherein
said compressor inlet passageway comprises a radially inwardly and
axially extending passage, said compressor outlet passageway
comprises a radially outwardly extending passageway, said first and
second turbine inlet passageways comprise radially inwardly
extending passages communicating with said vanes, and said
discharge passageways comprise axially extending passages directed
in opposite directions away from said compressor.
9. The compressor-turbine assembly as claimed in claim 8 wherein
said compressor impeller includes blades having axially extending
inlet portions and radially extending discharge portions, and said
first and second turbine impellers include blades having radially
extending inlet portions and axially extending discharge
portions.
10. A compressor-turbine assembly comprising:
a. a rotatable shaft;
b. a compressor impeller secured to said shaft;
c. casing means forming a compression chamber surrounding said
compressor impeller and forming inlet and discharge passageways
communicating with said compression chamber;
d. a first turbine impeller secured to said shaft axially spaced on
one side of said compressor impeller;
e. casing means forming an expansion chamber surrounding said first
turbine impeller and forming inlet and discharge passageways
communicating with said first expansion chamber;
f. a second turbine impeller secured to said shaft axially spaced
on the other side of said compressor impeller;
g. casing means forming an expansion chamber surrounding said
second turbine impeller and forming inlet and discharge passageways
communicating with said second expansion chamber;
h. said compressor inlet passageway comprises a radially inwardly
and axially extending passage, said compressor outlet passageway
comprises a radially outwardly extending passageway, said first and
second turbine inlet passageways comprise radially inwardly
extending passages communicating with said turbine impellers, and
said discharge passageways comprise axially extending passages
directed in opposite directions away from said compressor.
Description
BACKGROUND OF THE INVENTION
This invention relates to improvements on refrigeration producing
methods and apparatus particularly of the type employing a
multi-stage centrifugal compressor and a plurality of work
expansion engines of the turbine type.
It is known that the energy produced by work expansion of a
pressurized gas in a refrigeration process may be used to provide a
part of the power requirements of the process and it has been
proposed in the past to directly couple the work expansion machine
to a compressor which pressurizes the refrigeration gas. When the
energy produced by the work expansion is employed to drive a stage
of compression having power requirements substantially less than
the energy of work expansion, it is possible to design the work
expansion engine to operate at such a speed so as to obtain
economically acceptable efficiency for existing parameters
including pressure and temperature of the inlet gas and pressure
and temperature of the effluent. It is known that the critical
speed requirement of work expansion engines of the turbine type may
be materially relieved by effecting the work expansion in two
expansion turbines operating in series relation, that is, the
effluent of the first expansion turbine being fed to the inlet of
the second expansion turbine, and it has been proposed in the past
to directly couple a pair of expansion turbines operating in series
relation in the process to the final stage of the refrigerant gas
compressor. In such an arrangement, the series operation of the
expansion turbines makes it possible for the final compression
stage to operate in a region of higher efficiency and thus utilize
a greater percentage of the horsepower available from the work
expansion as compared to the processes in which a single expansion
turbine is coupled to a compression stage.
In another known refrigeration process, a pair of work expansion
turbines operate in parallel relation and at materially different
temperature levels; the turbines being fed with pressurized gas at
substantially the same pressure and discharge the gas at
substantially the same relatively low pressure. In such process,
the critical speed requirements of each expansion turbine is in no
way relieved, as in the process employing expansion turbines
operating in series as discusses above, and it has not been
possible heretofore to directly utilize at a high order of
efficiency the energy developed from both parallel operating
expansion turbines to drive the final stage of compression while
maintaining high expander efficiency.
The present invention provides a novel method and apparatus for
producing refrigeration employing multiple stages of refrigerant
compression and two stages of refrigerant work expansion in which
the horsepower developed by the work expansion stages is utilized
to drive the final stage of refrigerant compression. The novel
process employs and the novel apparatus includes a novel
expander-compressor-expander combination in which both expansion
turbines operate at different temperature levels under a
substantially equal pressure ratio and at high efficiency and in
which the power produced from both expansion turbines is applied
directly to the compressor operating as the final compression stage
and at a high order of efficiency to thus usefully employ the total
energy output of the expansion turbines within the normal range of
efficiency obtained for low temperature machinery. The novel
arrangement makes it possible to accomplish a greater percentage of
the total work of compression in the final compression stage and
more efficient compression of the gaseous refrigerant.
Other objects and advantages of the present invention will appear
from the following detailed description when considered in
connection with the accompanying drawings which disclose a
preferred embodiment of the invention. It is to be expressly
understood, however, that the drawings are designed for purposes of
illustration only and not as a definition of the limits of the
invention, reference for the latter purpose being had to the
appended claims.
DESCRIPTION OF THE DRAWINGS
In the drawings, in which similar elements are identified by
corresponding reference characters throughout the several
views:
FIG. 1 is a diagrammatic presentation of a refrigeration cycle
embodying the principles of the present invention;
FIG. 2 is a diagrammatic view, partly in section, of a novel
expander-compressor-expander unit provided by the present
invention;
FIG. 3 is a view in section of a part of the structure shown in
FIG. 2; and
FIG. 4 is a three-dimensional view of another part of the structure
shown in FIG. 2.
DESCRIPTION OF THE PREFERRED EMBODIMENT
The refrigeration cycle shown in FIG. 1 includes a novel
expander-compressor-expander unit 10 which includes a compressor 11
of the centrifugal type and centripetal turbine expanders 12 and 13
connected to a common shaft 14 with the compressor 11. The
compressor 11 comprises the final compression stage of the
refrigeration cycle which also includes a first stage compressor
15, driven by any suitable prime mover 16. The compressor 15
compresses refrigerant gas delivered to its suction inlet by
conduit 17 to an intermediate pressure and refrigerant gas at the
intermediate pressure is conducted by conduit 18 to the suction
inlet of the compressor 11 from which the refrigerant gas is
delivered in conduit 19 under relatively high pressure; the
compressors 11 and 15 may be provided with conventional after
coolers, not shown. The high pressure refrigerant gas is divided
with a first part flowing through passageway 20 of heat exchange
device 21 and a second part flowing through passageway 22 of heat
exchange device 23. The first part of the high pressure refrigerant
gas is cooled to below ambient temperature in the heat exchange
device 21 by heat interchange with an auxiliary refrigerant
supplied by conduits 24 and then conducted by conduit 25 to the
inlet of the turbine expander 12 wherein the refrigerant gas is
expanded with the production of external work to a pressure
corresponding substantially to the inlet pressure of the compressor
15 with concomitant cooling of the gas. The effluent of the turbine
expander 12 is conducted by conduits 26 and 27 for flow through
passageway 28 of the heat exchange device 23 and thereby warmed to
ambient temperature and then conducted by conduit 29 to the conduit
17 for recycling to the compressor 15. The second part of the high
pressure refrigerant gas is cooled upon flowing through the
passageway 22 and is then divided with one part of the cooled high
pressure refrigerant gas flowing through conduit 30 to the inlet of
the turbine expander 13 and with a second part flowing through
passageway 31 of the heat exchange device 32. In the turbine
expander 13, the cooled high pressure refrigerant gas is expanded
with work to a pressure slightly above the inlet pressure of the
compressor 15 with concomitant further cooling to a relatively low
temperature. The effluent of the turbine expander 13 may be
utilized to provide refrigeration externally of the cycle or
refrigeration for the cycle, or both. As shown, effluent of the
turbine expander 13 may be passed by conduit 33 to conduits 34 and
35, provided with control valves 36 and 37, respectively, for flow
through passageways 38 and 39, respectively, of heat exchange
device 40, in countercurrent heat interchange with an external
fluid flowing through passageway 41. The warm refrigerant gas
leaving the warm end of the passageway 39 is conducted by conduit
42 to the conduit 17 for flow to the inlet of the compressor 15,
while the refrigerant gas leaving the passageway 38 at a lower
temperature is conducted by conduit 43 to the conduit 27 for flow
through the passageway 28 and, hence, on to the inlet of the
compressor 15. Also, all or part of the effluent may be passed
through control valve 44 and conduit 45 for flow through passageway
46 of the heat exchange device 32 to provide refrigeration for the
cycle as described below.
The second part of the cooled high pressure refrigerant gas is
further cooled upon flowing through the passageway 31 of the heat
exchange device 32 and then expanded in valve 47 to a pressure
slightly above the suction pressure of the compressor 15 to effect
its partial liquefaction, and then fed by conduit 48 to phase
separator 49 where the liquefied refrigerant collects in a pool 50.
The liquefied refrigerant may be employed to cool an external fluid
flowed by conduit 51 having a control valve 52 through coil 53
immersed in the pool 50. Also, when the cycle is employed as a
liquefier, liquefied refrigerant may be withdrawn from the phase
separator through conduit 54 having a control valve 55. Gaseous
refrigerant comprising unliquefied refrigerant gas fed to the phase
separator and liquid refrigerant that may be vaporized in the phase
separator is withdrawn from the phase separator through conduit 56
and merged with the refrigerant gas in the conduit 45 for flow
through the passageways 46 and 28. Makeup refrigerant gas for the
cycle, as may be required to compensate for liquid refrigerant
withdrawn as product, for example, is fed to the suction inlet of
the compressor 15 by conduit 57 and, when the makeup refrigerant
gas is available at a pressure lower than suction pressure of the
compressor 15, the makeup refrigerant gas is fed by conduit 58
through a suitable compressor 59 to the conduit 57.
The power developed by the work expansion of the high temperature
high pressure refrigerant gas in the turbine expander 12 and the
power developed by the work expansion of the low temperature high
pressure refrigerant gas in the turbine expander 13 is applied
directly through the common shaft 14 to the compressor 11 to
provide the sole power source for the final compressor stage. As
described below, the expander-compressor-expander unit 10 is
characterized so that the expansion turbines 12 and 13, although
rotating at the same speed and notwithstanding the material
difference between their operating temperatures, both operate at a
high order of efficiency to satisfy fully the work expansion
refrigeration producing requirements of the cycle and both develop
a high percentage of the potential power of the work expansion
which is utilized to drive the final stage of compression which
makes possible, although the final compression stage is rotating at
the same speed as the expansion turbines, compression of the
refrigerant gas to the required high pressure from a lower
intermediate pressure thereby reducing the requirements of the
first stage of compression and achieving a reduction in capital
expenditure and operating power. It will be appreciated that the
flow of high pressure refrigerant gas discharged from the
compressor 11 is sufficient to satisfy the required flow of
refrigerant gas to the expansion turbines and also provides high
pressure refrigerant gas that is partially liquefied upon expansion
in valve 47.
As shown in FIG. 2, the expander-compressor-expander unit 10
includes a casing 60, which may be made up of sections bolted or
otherwise removably secured together, forming a housing for the
compressor 11, the turbine expander 12 and the turbine expander 13.
A shaft 61 is rotatably supported within the casing 60 by
journal-thrust bearings 62 and 63, preferably of the tilting pad
type, spaced axially of the shaft 61 and located inwardly of shaft
ends 64 and 65. An impeller 66 of the compressor 11 is mounted on
the shaft 61 for rotation therewith at a point intermediate the
bearings 62 and 63 and the impeller 66 includes radial blades 67
which are of constant radius at its input. Internal walls 68 of the
casing provide a chamber 69 within which the impeller 66 rotates
and an inlet passageway 70 and an outlet passageway 71
communicating with the chamber 69. The inlet passageway 70 extends
circumferentially about the shaft 61 and includes a portion 72
which extends from the outer regions of the casing 60, where it
communicates with fluid inlet 73, radially toward the shaft 61 and
a portion 74 which extends axially of the shaft and merges with the
chamber 69 at the inlet of the impeller 66. The outlet passageway
71 extends circumferentially about the shaft 61 radially outwardly
from communication with the chamber 69 at the discharge of the
impeller 66. The passageway 71 functions as a diffuser which may be
of the vaneless type and discharges into a volute 75 which
communicates with discharge outlet 76. The turbine expander 12
includes an impeller 77 mounted on and secured to the shaft end 65
and the turbine expander 13 includes an impeller 78 mounted on and
secured to the shaft end 64. The turbine expanders 12 and 13 are of
the radial-inflow or centripetal type and the impellers 77 and 78
are provided with radial blades 79, preferably of the open type,
having an axial dimension at the periphery of the impellers
corresponding substantially to the width of the nozzle inlet and a
substantially greater dimension, measured in a plane perpendicular
to the axis of rotation, at the exit of the impeller.
As shown in FIG. 3, internal walls 80 of the casing 60 define a
chamber 81 within which the impeller 77 rotates, the chamber 81
being shaped to conform to the outer configuration of the blades
79. The internal walls 80 also define a fluid inlet passageway 82
and an exhaust passageway 83 which diverges in the direction of
flow and is of circular cross section concentric with the axis of
rotation of the impeller. The inlet passageway 82 extends
circumferentially about the impeller 77 and communicates with the
chamber 81 at the inlet 84 of the blades 79, and extends radially
outwardly from the shaft into communication with a
circumferentially extending chamber 85, also defined by internal
walls 80 of the casing, into which high pressure gas to be expanded
is fed by input conduit 86. A plurality of spaced vanes 90 are
positioned in the passageway 82, in equally spaced relationship
about the circumference of the impeller, to provide therebetween
nozzles for directing the high pressure gas into the peripheral
entries of the blades 79. The vanes 90 are spaced radially from the
peripheral edge 84 of the blades 79 to permit the high pressure gas
discharged from the nozzles to follow a path influenced by the
position of the vanes 90 and the vanes 90 are adjustable in unison
to vary the mass and the direction of flow of the gas into the
impeller.
As shown more clearly in FIG. 4, each of the vanes 90 is mounted
for pivotal movement about an axis parallel to the axis of rotation
of the impeller by a shaft 93 supported in a wall 94 of the casing.
On the side of the wall 94 opposite the passageway 82, the shaft 93
is connected to an arm 95 provided with a slot 96 which receives a
pin 97 carried by a ring 98 positioned in concentric relation about
the axis of rotation of the impeller and rotatably supported in a
circular groove 99 of rectangular cross-section formed in the wall
94. A mechanism is provided for rotating the ring 98 relative to
the wall 94 to simultaneously rotate each of the vanes 90 in the
same direction and in the same degree about their respective
supporting shafts 93. Such mechanism includes a crank arm 100
connected outside of the casing 60 to an end of shaft 101 journaled
in the casing, the other end of the shaft 101 being rigidly
connected within the casing to one end of a crank arm 102. The
other end of the crank arm 102 is rigidly connected to shaft 103
secured to a rectangular block 104 received in a rectangular
opening 105 formed in the ring 98. With this arrangement, movement
of the crank arm 100 in one direction or the other effects rotation
of the ring 98 in a clockwise or counterclockwise direction, as the
case may be, relative to the wall 94 which in turn effects
simultaneous rotation of the vanes 90 in a clockwise or
counterclockwise direction about their supporting shafts 93.
The turbine expander 13 is constructed in a manner similar to the
turbine expander 12 but is differently dimensioned as described
below. The impeller 78 is a mirror image of the impeller 77 and
rotates within a chamber 110 defined by internal walls 111 of the
casing. The internal walls 111 also define a divergent discharge
passageway 112 and an inlet passageway 113, the latter passageway
communicating with an inlet chamber 114 to which high pressure gas
is supplied through input conduit 115. Adjustable vanes 116 are
located in the passageway 113 in a manner similar to the vanes 90
and a mechanism similar to the arrangement shown in FIG. 4 is
provided to effect adjustment of the vanes upon movement of crank
arm 100' connected to shaft 101'.
Seals 120, 121, 122, and 123, of the labyrinth type, are provided
about the shaft 61. Seals 120 and 121 are located inwardly of the
bearings 62, 63 on opposite sides of the compressor impeller 66,
and function to impede the flow of high pressure gas along the
shaft in a direction toward the turbine expanders, while seals 122
and 123, located between the bearing 62 and the impeller 78 and
between the bearing 63 and the impeller 77, respectively, function
to impede the flow of cold gas along the shaft inwardly toward the
compressor impeller. In order to prevent the flow of cold gas
inwardly toward the compressor impeller, high pressure gas is fed
to the seals 122 and 123 by passageways 124 and 125 leading from
the seals 120 and 121 to the seals 122 and 123, respectively.
In operation of turbine expanders of the foregoing type, the
incoming compressed gas is accelerated and directed by the nozzles
and enters the peripheral inlet of the radial blades with a large
tangential velocity and a small radial velocity and, upon flowing
through the impeller, the kinetic energy of the gas is transferred
to the impeller to effect its rotation and the gas leaves the
blades and enters the discharge passageway at a relatively low
velocity. For efficient operation, the tangential velocity of the
incoming gas and the angular velocity of the tips 84 of the radial
blades should correspond within narrow limits. Hence, the maximum
permissible tip speed of the blades is one of the controlling
parameters for expansion turbines operating at cryogenic
temperatures. The maximum angular velocity of the peripheral tips
of the blades is usually selected to be equal to 0.8 of Mach. I.
Within limits, the optimum speed of the tips of the blades for
given flow of a specific gas at a given pressure and temperature
may be established by correlating the speed of rotation of the
impeller and the diameter of the blades at the inlet of the
impeller. The provision of adjustable vanes, such as the adjustable
vanes 90 and 116 of the turbine expanders 12 and 13, makes it
possible to vary the mass and the direction of the high velocity
gas discharged from the nozzles which, together with the provision
of a radial space between the periphery of the blades and the
discharge end of the nozzles, makes it possible to operate the
expansion turbines at the designed efficiency over a wide range of
different input flow rates for a specific gas at a given pressure
and temperature.
Notwithstanding the limitations imposed on the designing of a low
temperature expansion turbine, it has been discovered that two
expansion turbines and a compressor may be designed as a unitary
structure with all of the impellers mounted on a common shaft and
rotating at a common speed, with the expansion turbines and the
compressor operating at high efficiency and with substantially the
total potential horsepower of the expansion being developed by the
expansion turbines and utilized as the sole power source for the
compressor which delivers the pressurized gas to both expansion
turbines and excess pressurized gas for liquefaction. The foregoing
has been accomplished, at least in part, by operating the expansion
turbines at a constant input pressure or at a substantially equal
pressure ratio, by operating the expansion turbines at
substantially different temperature levels, by maintaining
substantially equal flow to each of the expansion turbines, and by
utilizing a constant input compressor. The feature of rotating the
compressor and both expansion turbines at the same speed permits
the impeller of the compressor and the impellers of the expansion
turbines to be mounted on a common shaft to provide a compact
assembly which offers special advantages in low temperature
equipment especially in view of the high rotating speeds involved.
The further feature provided by the present invention of mounting
the impellers of the two expansion turbines on opposite sides of
the compressor impeller aids in balancing the machine and permits
counter application of the thrust developed by the expansion
turbine impellers and partial compensation of the thrust developed
by the compressor impeller.
An expander-compressor-expander unit constructed in accordance with
the principles of the present invention has been operated for
extended periods of time in a refrigeration cycle of the type shown
in FIG. 1. The compressor was of the centrifugal type and included
an open impeller with the radial blades at the periphery having a
diameter of about 5.00 inches, the inlet nozzle and the diffuser
being provided without vanes. The turbine expanders were of the
centripetal type having impellers provided with open radial blades;
the impeller of the high temperature turbine expander having a
diameter at its inlet of 5.25 inches and the impeller of the low
temperature turbine expander having a diameter of 4.50 inches at
its inlet. The turbine expanders were provided with adjustable
nozzles, as described above, and the nozzle control arms were
ganged together for simultaneous control; however, means were
provided for independent adjustment of the nozzles of each turbine
expander. The following table provides data on the performance of
the expander-compressor-expander unit utilizing nitrogen as the
refrigerant, the data being collected at different times during a
period of continuous operation which exceeded 5 days:
Time 1 Time 2 Time 3 Time 4
__________________________________________________________________________
Compressor Inlet pressure, psig 400 405 405 400 Inlet Temperature,
.degree.F. 81 83 82 80 Discharge pressure, psig 545 555 555 550
Discharge temperature, .degree.F. 155 157 157 157 Flow, standard
cubic 1920 1900 1900 1920 feet/hour .times. 1000 Adiabatic
efficiency, % 63.5 64.8 64 63 Horsepower 1025 1020 1030 1070
High Temperature Expansion Turbine Inlet pressure, psig 525 535 535
530 Inlet temperature, .degree.F. 17 21 22 21 Discharge pressure,
psig 69 70 70 68 Discharge temperature, .degree.F. -120 - 123 -122
-124 Flow, standard cubic 566 520 520 520 feet/hour .times. 1000
Adiabatic efficiency, % 66.5 67.5 68. 6 69.3 Horsepower 499 472 480
488
Low Temperature Expansion Turbine Inlet pressure, psig 520 530 530
525 Inlet temperature, .degree.F. -178 - 171 -173 - 159 Discharge
pressure, psig 75 71 71 70 Discharge temperature, .degree.F. -282 -
283 -283 - 272 Flow, standard cubic 1075 990 990 990 feet/hour
.times. 1000 Adiabatic efficiency, % 80.2 86.0 85.0 81.0 Horsepower
553 585 575 596 Shaft speed, rpm 35,500 36,000 35,500 36,000
__________________________________________________________________________
During the foregoing operation, the cycle of FIG. 1 functioned as a
closed system to provide sources of refrigeration for a low
pressure air separation system to increase liquid producing
capacity. Compressed air to be separated flowed through the
passageway 41 of the heat exchange device 40 and the liquid-vapor
nitrogen mixture in conduit 48 was introduced into the high
pressure fractionating column and an equivalent mass of nitrogen
vapor was withdrawn from the fractionating column by conduit 56.
Considering the data collected at Time 1 of the above table as
being typical, the nitrogen gas was discharged from the compressor
11 at about 100.degree.F. and under a pressure of about 545
p.s.i.g. and was divided with about 30 percent flowing through
passageway 20 of heat exchange device 21 and with about 70 percent
flowing through passageway 22 of heat exchange device 23. The
compressed nitrogen gas was cooled in heat exchange device 21 to
about 17.degree.F. in heat interchange with a freon refrigerant and
was introduced into the inlet of the turbine expander 12 under a
pressure of about 525 p.s.i.g. The major portion of the nitrogen
gas was cooled in heat exchange device 23 to about -178.degree.F.
and then divided with about 80 percent being fed to the inlet of
the turbine expander 13 under a pressure of about 520 p.s.i.g. The
remaining 20 percent of the high pressure nitrogen gas was cooled
to about -276.degree.F. upon flowing through the passageway 31 of
the heat exchange device 32, then expanded in valve 47 to about 75
p.s.i.g. and then introduced into the fractionating column (phase
separator 49) at a temperature of about -285.degree.F., partially
in liquid phase. The effluent of the expansion turbine 13 at a
temperature of about -282.degree.F. and under a pressure of about
75 p.s.i.g. was subdivided with about 52 percent flowing through
the conduit 45, about 28 percent flowing through the passageway 38
of the heat exchange device 40 and about 20 percent flowing through
the passageway 39 of such heat exchange device. The cold nitrogen
gas in the conduit 45 was merged with the nitrogen vapor at a
temperature of -285.degree.F. withdrawn from the phase separator
49, and the merged stream was warmed upon flowing through the
passageway 46 of the heat exchange device 32 to about
-179.degree.F. and then combined with the nitrogen gas withdrawn
from the passageway 38 at about -164.degree.F.; the combined
streams being warmed to about 75.degree.F. upon flowing through the
passageway 28 of the heat exchange device 23 and then returned to
the inlet of the compressor 15 by way of the conduit 17. The
nitrogen gas flowing through the passageway 39 of the heat exchange
device 40 was withdrawn from that passageway at about 84.degree.F.
and conducted by conduit 42 to the conduit 17 onto the inlet of the
compressor 15. The pressurized nitrogen gas from the compressor 15,
at about 80.degree.F. and under a pressure of about 400 p.s.i.g.,
was fed by conduit 18 to the inlet of the compressor 11.
Although only one embodiment of the present invention has been
disclosed and described herein, it is to be expressly understood
that various changes and modifications may be made therein without
departing from the spirit of the invention as well understood by
those skilled in the art. Reference therefore will be had to the
appended claims for a definition of the limits of the
invention.
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