U.S. patent application number 17/466602 was filed with the patent office on 2021-12-23 for compressor with liquid injection cooling.
The applicant listed for this patent is HICOR TECHNOLOGIES, INC.. Invention is credited to Phil NELSON, Jeremy PITTS, John WALTON.
Application Number | 20210396232 17/466602 |
Document ID | / |
Family ID | 1000005822567 |
Filed Date | 2021-12-23 |
United States Patent
Application |
20210396232 |
Kind Code |
A1 |
WALTON; John ; et
al. |
December 23, 2021 |
COMPRESSOR WITH LIQUID INJECTION COOLING
Abstract
A compressor includes: a casing with an inner wall defining a
compression chamber, an inlet leading into the compression chamber,
and an outlet leading out of the compression chamber; a rotor
rotatably coupled to the casing for rotation relative to the
casing; and a gate coupled to the casing for movement relative to
the casing. The gate may be pivotally, or translationally coupled
to the casing. A hydrostatic bearing may be disposed between the
gate and casing. A plurality of compressors may be mechanically
linked together such that their compression cycles are out of
phase.
Inventors: |
WALTON; John; (Portland,
ME) ; NELSON; Phil; (Houston, TX) ; PITTS;
Jeremy; (Solon, OH) |
|
Applicant: |
Name |
City |
State |
Country |
Type |
HICOR TECHNOLOGIES, INC. |
Houston |
TX |
US |
|
|
Family ID: |
1000005822567 |
Appl. No.: |
17/466602 |
Filed: |
September 3, 2021 |
Related U.S. Patent Documents
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Application
Number |
Filing Date |
Patent Number |
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16566657 |
Sep 10, 2019 |
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17466602 |
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15563061 |
Sep 29, 2017 |
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PCT/US2016/024803 |
Mar 29, 2016 |
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16566657 |
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62139884 |
Mar 30, 2015 |
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Current U.S.
Class: |
1/1 |
Current CPC
Class: |
F01C 21/0809 20130101;
F04C 27/001 20130101; F04C 27/02 20130101; F04C 29/0007 20130101;
F04C 18/46 20130101; F04C 2240/54 20130101 |
International
Class: |
F04C 29/00 20060101
F04C029/00; F04C 18/46 20060101 F04C018/46; F01C 21/08 20060101
F01C021/08; F04C 27/00 20060101 F04C027/00; F04C 27/02 20060101
F04C027/02 |
Claims
1-20. (canceled)
21. A compressor system comprising: a plurality of sub-compressors,
each sub-compressor comprising: a wall defining a compression
chamber, a fluid inlet leading into the compression chamber, a
fluid outlet leading out of the compression chamber, a rotor
rotatably coupled to the wall for rotation relative to the wall,
and a gate coupled to the wall for movement relative to the wall,
the gate comprising a sealing edge, the gate being operable to move
relative to the wall to locate the sealing edge proximate to the
rotor as the rotor rotates such that the gate separates an inlet
volume and a compression volume in the compression chamber, the
inlet and outlet being disposed on opposite sides of the sealing
edge from each other; and a mechanical linkage between the rotors
of the plurality of sub-compressors, the mechanical linkage
connecting between the rotors of the plurality of sub-compressors
such that compression cycles of the plurality of sub-compressors
are out of phase with each other.
22. The compressor system of claim 21, wherein the plurality of
sub-compressors comprises n sub-compressors, and wherein the
mechanical linkage connects the rotors such that the compression
cycle of each of the n sub-compressors is out of phase with
phase-wise adjacent ones of the n sub-compressors by 360/n degrees,
and wherein 2.ltoreq.n.ltoreq.100.
23. The compressor system of claim 21, wherein the mechanical
linkage comprises a drive shaft that extends through each of the
plurality of sub-compressors and is coupled to the rotors of each
of the plurality of sub-compressors for common rotation with the
rotors of each of the plurality of sub-compressors relative to the
walls of each of the plurality of sub-compressors.
33-36. (canceled)
37. The compressor system of claim 21, wherein each of the
plurality of sub-compressors comprises a respective casing that
defines the wall of the respective sub-compressor.
38. The compressor system of claim 21, further comprising a casing
that defines each of the walls and compression chambers of each of
the sub-compressors.
39. The compressor system of claim 38, wherein the mechanical
linkage comprises a drive shaft rotatably mounted to the casing for
rotation relative to the casing, wherein the rotors of each of the
plurality of sub-compressors are mounted to the drive shaft for
common rotation with the drive shaft relative to the casing.
40. The compressor system of claim 21, further comprising an inlet
manifold that fluidly interconnects the inlets of each of the
plurality of sub-compressors.
41. The compressor system of claim 41, wherein each of the
sub-compressors comprises a check valve disposed in the inlet of
the respective sub-compressor downstream from where the inlets of
the plurality of sub-compressors diverge from each other.
42. The compressor system of claim 21, further comprising a
discharge manifold that fluidly interconnects the outlets of each
of the plurality of sub-compressors.
43. The compressor system of claim 42, wherein each of the
sub-compressors comprises a check valve disposed in the outlet of
the respective sub-compressor upstream from where the outlets of
the plurality of sub-compressors interconnect.
Description
CROSS-REFERENCE TO RELATED APPLICATION
[0001] This application claims the benefit of U.S. Provisional
Application Ser. No. 62/139,884, filed on Mar. 30, 2015, the
content of which is hereby incorporated herein by reference in its
entirety.
BACKGROUND
1. Technical Field
[0002] The invention generally relates to fluid pumps, such as
compressors and expanders.
2. Related Art
[0003] Compressors have typically been used for a variety of
applications, such as air compression, vapor compression for
refrigeration, and compression of industrial gases. Compressors can
be split into two main groups, positive displacement and dynamic.
Positive displacement compressors reduce the compression volume in
the compression chamber to increase the pressure of the fluid in
the chamber. This is done by applying force to a drive shaft that
is driving the compression process. Dynamic compressors work by
transferring energy from a moving set of blades to the working
fluid.
[0004] Positive displacement compressors can take a variety of
forms. They are typically classified as reciprocating or rotary
compressors. Reciprocating compressors are commonly used in
industrial applications where higher pressure ratios are necessary.
They can easily be combined into multistage machines, although
single stage reciprocating compressors are not typically used at
pressures above 80 psig. Reciprocating compressors use a piston to
compress the vapor, air, or gas, and have a large number of
components to help translate the rotation of the drive shaft into
the reciprocating motion used for compression. This can lead to
increased cost and reduced reliability. Reciprocating compressors
also suffer from high levels of vibration and noise. This
technology has been used for many industrial applications such as
natural gas compression.
[0005] Rotary compressors use a rotating component to perform
compression. As noted in the art, rotary compressors typically have
the following features in common: (1) they impart energy to the gas
being compressed by way of an input shaft moving a single or
multiple rotating elements; (2) they perform the compression in an
intermittent mode; and (3) they do not use inlet or discharge
valves. (Brown, Compressors: Selection and Sizing, 3rd Ed., at 6).
As further noted in Brown, rotary compressor designs are generally
suitable for designs in which less than 20:1 pressure ratios and
1000 CFM flow rates are desired. For pressure ratios above 20:1,
Royce suggests that multistage reciprocating compressors should be
used instead.
[0006] Typical rotary compressor designs include the rolling
piston, screw compressor, scroll compressor, lobe, liquid ring, and
rotary vane compressors. Each of these traditional compressors has
deficiencies for producing high pressure, near isothermal
conditions.
[0007] The design of a rotating element/rotor/lobe against a
radially moving element/piston to progressively reduce the volume
of a fluid has been utilized as early as the mid-19th century with
the introduction of the "Yule Rotary Steam Engine." Developments
have been made to small-sized compressors utilizing this
methodology into refrigeration compression applications. However,
current Yule-type designs are limited due to problems with
mechanical spring durability (returning the piston element) as well
as chatter (insufficient acceleration of the piston in order to
maintain contact with the rotor).
[0008] For commercial applications, such as compressors for
refrigerators, small rolling piston or rotary vane designs are
typically used. (P N Ananthanarayanan, Basic Refrigeration and Air
Conditioning, 3rd Ed., at 171-72.) In these designs, a closed
oil-lubricating system is typically used.
[0009] Rolling piston designs typically allow for a significant
amount of leakage between an eccentrically mounted circular rotor,
the interior wall of the casing, and/or the vane that contacts the
rotor. By spinning the rolling piston faster, the leakages are
deemed acceptable because the desired pressure and flow rate for
the application can be easily reached even with these losses. The
benefit of a small self-contained compressor is more important than
seeking higher pressure ratios.
[0010] Rotary vane designs typically use a single circular rotor
mounted eccentrically in a cylinder slightly larger than the rotor.
Multiple vanes are positioned in slots in the rotor and are kept in
contact with the cylinder as the rotor turns typically by spring or
centrifugal force inside the rotor. The design and operation of
these type of compressors may be found in Mark's Standard Handbook
for Mechanical Engineers, Eleventh Edition, at 14:33-34.
[0011] In a sliding-vane compressor design, vanes are mounted
inside the rotor to slide against the casing wall. Alternatively,
rolling piston designs utilize a vane mounted within the cylinder
that slides against the rotor. These designs are limited by the
amount of restoring force that can be provided and thus the
pressure that can be yielded.
[0012] Each of these types of prior art compressors has limits on
the maximum pressure differential that it can provide. Typical
factors include mechanical stresses and temperature rise. One
proposed solution is to use multistaging. In multistaging, multiple
compression stages are applied sequentially. Intercooling, or
cooling between stages, is used to cool the working fluid down to
an acceptable level to be input into the next stage of compression.
This is typically done by passing the working fluid through a heat
exchanger in thermal communication with a cooler fluid. However,
intercooling can result in some condensation of liquid and
typically requires filtering out of the liquid elements.
Multistaging greatly increases the complexity of the overall
compression system and adds costs due to the increased number of
components required. Additionally, the increased number of
components leads to decreased reliability and the overall size and
weight of the system are markedly increased.
[0013] For industrial applications, single- and double-acting
reciprocating compressors and helical-screw type rotary compressors
are most commonly used. Single-acting reciprocating compressors are
similar to an automotive type piston with compression occurring on
the top side of the piston during each revolution of the
crankshaft. These machines can operate with a single-stage
discharging between 25 and 125 psig or in two stages, with outputs
ranging from 125 to 175 psig or higher. Single-acting reciprocating
compressors are rarely seen in sizes above 25 HP. These types of
compressors are typically affected by vibration and mechanical
stress and require frequent maintenance. They also suffer from low
efficiency due to insufficient cooling.
[0014] Double-acting reciprocating compressors use both sides of
the piston for compression, effectively doubling the machine's
capacity for a given cylinder size. They can operate as a
single-stage or with multiple stages and are typically sized
greater than 10 HP with discharge pressures above 50 psig. Machines
of this type with only one or two cylinders require large
foundations due to the unbalanced reciprocating forces.
Double-acting reciprocating compressors tend to be quite robust and
reliable, but are not sufficiently efficient, require frequent
valve maintenance, and have extremely high capital costs.
[0015] Lubricant-flooded rotary screw compressors operate by
forcing fluid between two intermeshing rotors within a housing
which has an inlet port at one end and a discharge port at the
other. Lubricant is injected into the chamber to lubricate the
rotors and bearings, take away the heat of compression, and help to
seal the clearances between the two rotors and between the rotors
and housing. This style of compressor is reliable with few moving
parts. However, it becomes quite inefficient at higher discharge
pressures (above approximately 200 psig) due to the intermeshing
rotor geometry being forced apart and leakage occurring. In
addition, lack of valves and a built-in pressure ratio leads to
frequent over or under compression, which translates into
significant energy efficiency losses.
[0016] Rotary screw compressors are also available without
lubricant in the compression chamber, although these types of
machines are quite inefficient due to the lack of lubricant helping
to seal between the rotors. They are a requirement in some process
industries such as food and beverage, semiconductor, and
pharmaceuticals, which cannot tolerate any oil in the compressed
air used in their processes. Efficiency of dry rotary screw
compressors are 15-20% below comparable injected lubricated rotary
screw compressors and are typically used for discharge pressures
below 150 psig.
[0017] Using cooling in a compressor is understood to improve upon
the efficiency of the compression process by extracting heat,
allowing most of the energy to be transmitted to the gas and
compressing with minimal temperature increase. Liquid injection has
previously been utilized in other compression applications for
cooling purposes. Further, it has been suggested that smaller
droplet sizes of the injected liquid may provide additional
benefits.
[0018] In U.S. Pat. No. 4,497,185, lubricating oil was intercooled
and injected through an atomizing nozzle into the inlet of a rotary
screw compressor. In a similar fashion, U.S. Pat. No. 3,795,117
uses refrigerant, though not in an atomized fashion, that is
injected early in the compression stages of a rotary screw
compressor. Rotary vane compressors have also attempted finely
atomized liquid injection, as seen in U.S. Pat. No. 3,820,923.
[0019] Published International Pat. App. No. WO 2010/017199 and
U.S. Pat. Pub. No. 2011/0023814 relate to a rotary engine design
using a rotor, multiple gates to create the chambers necessary for
a combustion cycle, and an external cam-drive for the gates. The
force from the combustion cycle drives the rotor, which imparts
force to an external element. Engines are designed for a
temperature increase in the chamber and high temperatures
associated with the combustion that occurs within an engine.
Increased sealing requirements necessary for an effective
compressor design are unnecessary and difficult to achieve.
Combustion forces the use of positively contacting seals to achieve
near perfect sealing, while leaving wide tolerances for metal
expansion, taken up by the seals, in an engine. Further, injection
of liquids for cooling would be counterproductive and coalescence
is not addressed.
[0020] Liquid mist injection has been used in compressors, but with
limited effectiveness. In U.S. Pat. No. 5,024,588, a liquid
injection mist is described, but improved heat transfer is not
addressed. In U.S. Pat. Publication. No. U.S. 2011/0023977, liquid
is pumped through atomizing nozzles into a reciprocating piston
compressor's compression chamber prior to the start of compression.
It is specified that liquid will only be injected through atomizing
nozzles in low pressure applications. Liquid present in a
reciprocating piston compressor's cylinder causes a high risk for
catastrophic failure due to hydrolock, a consequence of the
incompressibility of liquids when they build up in clearance
volumes in a reciprocating piston, or other positive displacement,
compressor. To prevent hydrolock situations, reciprocating piston
compressors using liquid injection will typically have to operate
at very slow speeds, adversely affecting the performance of the
compressor.
[0021] U.S. Patent Application Publication No. 2013-0209299, titled
"Compressor With Liquid Injection Cooling" discloses another rotary
compressor with liquid injection cooling. The entire contents of
U.S. Patent Application Publication No. 2013-0209299 are
incorporated herein by reference in its entirety.
BRIEF SUMMARY
[0022] The presently preferred embodiments are directed to rotary
compressor designs. These designs are particularly suited for high
pressure applications, typically above 200 psig with pressure
ratios typically above that for existing high-pressure positive
displacement compressors.
[0023] One or more embodiments provides a compressor that includes:
a casing with an inner wall defining a compression chamber; a drive
shaft and rotor rotatably coupled to the casing for common rotation
relative to the casing, the rotor having a non-circular profile;
and a gate coupled to the casing for pivotal movement relative to
the casing, the gate comprising a sealing edge, the gate being
operable to move relative to the casing to locate the sealing edge
proximate to the rotor as the rotor rotates such that the gate
separates an inlet volume and a compression volume in the
compression chamber.
[0024] One or more embodiments provides a compressor that includes:
a casing with an inner wall defining a compression chamber, an
inlet leading into the compression chamber, and an outlet leading
out of the compression chamber; a drive shaft and rotor rotatably
coupled to the casing for common rotation relative to the casing,
the rotor having a non-circular profile; a gate coupled to the
casing for movement relative to the casing, the gate comprising a
sealing edge, the gate being operable to move relative to the
casing to locate the sealing edge proximate to the rotor as the
rotor rotates such that the gate separates an inlet volume and a
compression volume in the compression chamber, the inlet and outlet
being disposed on opposite sides of the sealing edge from each
other; and an outlet manifold in fluid communication with the
outlet, wherein the outlet is elongated in a direction parallel to
a rotational axis of the drive shaft, wherein the outlet manifold
defines an interior passageway, and wherein the passageway varies
in cross-sectional shape between an entrance into the manifold and
an exit out of the manifold, and wherein the outlet manifold
comprises a plurality of vanes disposed in the interior passageway
to direct the flow of working fluid through the outlet
manifold.
[0025] One or more embodiments provides a compressor that includes:
a casing with an inner wall defining a compression chamber, an
inlet leading into the compression chamber, and an outlet leading
out of the compression chamber; a rotor coupled to the casing for
rotation relative to the casing; a gate movably coupled to one of
the casing and rotor for movement relative to the one of the casing
and rotor, the gate comprising a sealing edge, the gate being
operable to locate the sealing edge proximate to the other of the
casing and rotor as the rotor rotates; and a hydrostatic bearing
arrangement disposed between (1) the gate and (2) the one of the
casing and rotor to reduce friction when the gate moves during
operation of the compressor.
[0026] One or more embodiments provides a compressor that includes:
a compression chamber casing with an inner wall defining a
compression chamber, an inlet leading into the compression chamber,
and an outlet leading out of the compression chamber; a drive shaft
and rotor rotatably coupled to the compression chamber casing for
common rotation relative to the compression chamber casing; a gate
coupled to the compression chamber casing for movement relative to
the compression chamber casing, the gate comprising a sealing edge,
the gate being operable to move relative to the compression chamber
casing to locate the sealing edge proximate to the rotor as the
rotor rotates such that the gate separates an inlet volume and a
compression volume in the compression chamber, the inlet and outlet
being disposed on opposite sides of the sealing edge from each
other; and a gate positioning system coupled to the gate, the gate
positioning system being shaped and configured to reciprocally move
the gate during rotation of the rotor so that the sealing edge
remains proximate to the rotor during rotation of the rotor.
[0027] According to various embodiments, the gate positioning
system includes a cam shaft rotatably coupled to the compression
chamber casing for rotation relative to the compression chamber
casing, the cam shaft being spaced from the drive shaft, the cam
shaft being connected to the drive shaft so as to be rotationally
driven by the drive shaft, a cam rotatably coupled to the
compression chamber casing for concentric rotation with the cam
shaft relative to the compression chamber casing, a cam follower
mounted to the gate for movement with the gate relative to the
compression chamber casing, the cam follower abutting the cam so
that rotation of the cam causes the cam follower and gate to move
relative to the compression chamber casing.
[0028] One or more embodiments provides a compressor system that
includes: a plurality of compressors. Each compressor may include a
casing with an inner wall defining a compression chamber, an inlet
leading into the compression chamber, and an outlet leading out of
the compression chamber, a rotor rotatably coupled to the casing
for rotation relative to the casing, and a gate coupled to the
casing for movement relative to the casing, the gate comprising a
sealing edge, the gate being operable to move relative to the
casing to locate the sealing edge proximate to the rotor as the
rotor rotates such that the gate separates an inlet volume and a
compression volume in the compression chamber, the inlet and outlet
being disposed on opposite sides of the sealing edge from each
other. The system includes a mechanical linkage between the rotors
of the plurality of compressors, the mechanical linkage connecting
between the rotors such that compression cycles of the plurality of
compressors are out of phase with each other.
[0029] One or more embodiments provides a compressor that includes:
a casing with an inner wall defining a compression chamber, an
inlet leading into the compression chamber, and an outlet leading
out of the compression chamber; a drive shaft and rotor rotatably
coupled to the casing for common rotation relative to the casing
such that when the rotor is rotated, the compressor compresses
working fluid that enters the compression chamber from the inlet,
and forces compressed working fluid out of the compression chamber
through the outlet; and a mechanical seal located at an interface
between the drive shaft and casing where the drive shaft passes
through the casing.
[0030] According to various embodiments, the mechanical seal
includes: first, second, and third seals disposed sequentially
along a leakage path between the drive shaft and casing rotor, a
source of pressurized hydraulic fluid, and a hydraulic fluid
passageway that connects the source to a space along the leakage
path between the second and third seals so as to keep the space
pressurized with hydraulic fluid.
[0031] One or more embodiments provides a non-circular seal for
sealing an interface between two moving parts. The seal includes a
non-circular structural base (e.g., comprising steel) having a
closed perimeter; and a low friction sealing material (e.g.,
graphite or Teflon) bonded to the base.
[0032] One or more embodiments provides a compressor that includes:
a casing with an inner wall defining a compression chamber, an
inlet leading into the compression chamber, and an outlet leading
out of the compression chamber; a rotor rotatably coupled to the
casing for rotation relative to the casing such that when the rotor
is rotated, the compressor compresses working fluid that enters the
compression chamber from the inlet, and forces compressed working
fluid out of the compression chamber through the outlet; a gate
coupled to the casing for reciprocating movement relative to the
casing, the gate comprising a sealing edge, the gate being operable
to move relative to the casing to locate the sealing edge proximate
to the rotor as the rotor rotates such that the gate separates an
inlet volume and a compression volume in the compression chamber;
and a mechanical seal located at an interface between the gate and
casing. The mechanical seal includes: first, second, and third
seals disposed sequentially along a leakage path between the gate
and casing, a source of pressurized hydraulic fluid, and a
hydraulic fluid passageway that connects the source to a space
along the leakage path between the second and third seals so as to
keep the space pressurized with hydraulic fluid.
[0033] According to various embodiments, the mechanical seal
further includes a vent disposed between the first and second
seals, the vent being fluidly connected to the inlet so as to
direct working fluid that leaks from the compression chamber past
the first seal back to the inlet.
[0034] According to various embodiments, the first, second, and
third seals are all supported by a removable housing, such that the
first, second, and third seals and housing can be installed into
the casing as a single unit.
[0035] According to various embodiments, the mechanical seal
comprises n sequential seals along the leakage path between the
gate and casing, wherein 3.ltoreq.n.ltoreq.50, wherein n includes
the first, second, and third seals, wherein one or more spaces
between adjacent ones of the seals are filled with pressurized
hydraulic fluid, and wherein one or more spaces between adjacent
ones of the seals comprise a vent that is fluidly connected on the
inlet.
[0036] These and other aspects of various non-limiting embodiments
of the present invention, as well as the methods of operation and
functions of the related elements of structure and the combination
of parts and economies of manufacture, will become more apparent
upon consideration of the following description and the appended
claims with reference to the accompanying drawings, all of which
form a part of this specification, wherein like reference numerals
designate corresponding parts in the various figures. In one
embodiment of the invention, the structural components illustrated
herein are drawn to scale. It is to be expressly understood,
however, that the drawings are for the purpose of illustration and
description only and are not intended as a definition of the limits
of the invention. In addition, it should be appreciated that
structural features shown or described in any one embodiment herein
can be used in other embodiments as well. As used in the
specification and in the claims, the singular form of "a", "an",
and "the" include plural referents unless the context clearly
dictates otherwise.
[0037] All closed-ended (e.g., between A and B) and open-ended
(greater than C) ranges of values disclosed herein explicitly
include all ranges that fall within or nest within such ranges. For
example, a disclosed range of 1-10 is understood as also
disclosing, among other ranged, 2-10, 1-9, 3-9, etc.
BRIEF DESCRIPTION OF THE DRAWINGS
[0038] Embodiments of the invention can be better understood with
reference to the following drawings and description. The components
in the figures are not necessarily to scale, emphasis instead being
placed upon illustrating the principles of various embodiments of
the invention. Moreover, in the figures, like referenced numerals
designate corresponding parts throughout the different views.
[0039] FIG. 1 is a perspective view of a rotary compressor with a
spring-backed cam drive in accordance with an embodiment of the
present invention.
[0040] FIG. 2 is a right-side view of a rotary compressor with a
spring-backed cam drive in accordance with an embodiment of the
present invention.
[0041] FIG. 3 is a left-side view of a rotary compressor with a
spring-backed cam drive in accordance with an embodiment of the
present invention.
[0042] FIG. 4 is a front view of a rotary compressor with a
spring-backed cam drive in accordance with an embodiment of the
present invention.
[0043] FIG. 5 is a back view of a rotary compressor with a
spring-backed cam drive in accordance with an embodiment of the
present invention.
[0044] FIG. 6 is a top view of a rotary compressor with a
spring-backed cam drive in accordance with an embodiment of the
present invention.
[0045] FIG. 7 is a bottom view of a rotary compressor with a
spring-backed cam drive in accordance with an embodiment of the
present invention.
[0046] FIG. 8 is a cross-sectional view of a rotary compressor with
a spring-backed cam drive in accordance with an embodiment of the
present invention.
[0047] FIG. 9 is a perspective view of rotary compressor with a
belt-driven, spring-biased gate positioning system in accordance
with an embodiment of the present invention.
[0048] FIG. 10 is a perspective view of a rotary compressor with a
dual cam follower gate positioning system in accordance with an
embodiment of the present invention.
[0049] FIG. 11 is a right-side view of a rotary compressor with a
dual cam follower gate positioning system in accordance with an
embodiment of the present invention.
[0050] FIG. 12 is a left-side view of a rotary compressor with a
dual cam follower gate positioning system in accordance with an
embodiment of the present invention.
[0051] FIG. 13 is a front view of a rotary compressor with a dual
cam follower gate positioning system in accordance with an
embodiment of the present invention.
[0052] FIG. 14 is a back view of a rotary compressor with a dual
cam follower gate positioning system in accordance with an
embodiment of the present invention.
[0053] FIG. 15 is a top view of a rotary compressor with a dual cam
follower gate positioning system in accordance with an embodiment
of the present invention.
[0054] FIG. 16 is a bottom view of a rotary compressor with a dual
cam follower gate positioning system in accordance with an
embodiment of the present invention.
[0055] FIG. 17 is a cross-sectional view of a rotary compressor
with a dual cam follower gate positioning system in accordance with
an embodiment of the present invention.
[0056] FIG. 18 is perspective view of a rotary compressor with a
belt-driven gate positioning system in accordance with an
embodiment of the present invention.
[0057] FIG. 19 is perspective view of a rotary compressor with an
offset gate guide positioning system in accordance with an
embodiment of the present invention.
[0058] FIG. 20 is a right-side view of a rotary compressor with an
offset gate guide positioning system in accordance with an
embodiment of the present invention.
[0059] FIG. 21 is a front view of a rotary compressor with an
offset gate guide positioning system in accordance with an
embodiment of the present invention.
[0060] FIG. 22 is a cross-sectional view of a rotary compressor
with an offset gate guide positioning system in accordance with an
embodiment of the present invention.
[0061] FIG. 23 is perspective view of a rotary compressor with a
linear actuator gate positioning system in accordance with an
embodiment of the present invention.
[0062] FIGS. 24A and B are right side and cross-section views,
respectively, of a rotary compressor with a magnetic drive gate
positioning system in accordance with an embodiment of the present
invention
[0063] FIG. 25 is perspective view of a rotary compressor with a
scotch yoke gate positioning system in accordance with an
embodiment of the present invention.
[0064] FIGS. 26A-F are cross-sectional views of the inside of an
embodiment of a rotary compressor with a contacting tip seal in a
compression cycle in accordance with an embodiment of the present
invention.
[0065] FIGS. 27A-F are cross-sectional views of the inside of an
embodiment of a rotary compressor without a contacting tip seal in
a compression cycle in accordance with another embodiment of the
present invention.
[0066] FIG. 28 is perspective, cross-sectional view of a rotary
compressor in accordance with an embodiment of the present
invention.
[0067] FIG. 29 is a left-side view of an additional liquid
injectors embodiment of the present invention.
[0068] FIG. 30 is a cross-section view of a rotor design in
accordance with an embodiment of the present invention.
[0069] FIGS. 31A-D are cross-sectional views of rotor designs in
accordance with various embodiments of the present invention.
[0070] FIGS. 32A and B are perspective and right-side views of a
drive shaft, rotor, and gate in accordance with an embodiment of
the present invention.
[0071] FIG. 33 is a perspective view of a gate with exhaust ports
in accordance with an embodiment of the present invention.
[0072] FIGS. 34A and B are a perspective view and magnified view of
a gate with notches, respectively, in accordance with an embodiment
of the present invention.
[0073] FIG. 35 is a cross-sectional, perspective view a gate with a
rolling tip in accordance with an embodiment of the present
invention.
[0074] FIG. 36 is a cross-sectional front view of a gate with a
liquid injection channel in accordance with an embodiment of the
present invention.
[0075] FIG. 37 is a graph of the pressure-volume curve achieved by
a compressor according to one or more embodiments of the present
invention relative to adiabatic and isothermal compression.
[0076] FIGS. 38A-38D show the sequential compression cycle and
liquid coolant injection locations, directions, and timing
according to one or more embodiments of the invention.
[0077] FIG. 39 is a perspective view of a compressor according to
an alternative embodiment.
[0078] FIG. 40 is a cross-sectional view of the compressor in FIG.
39, taken along an axis of the compressor's drive shaft.
[0079] FIG. 41 is an exploded view of the compressor in FIG.
39.
[0080] FIG. 42 is an end view of the compressor in FIG. 39.
[0081] FIG. 43 is a cross-sectional view of the compressor in FIG.
39, taken in a plane that is perpendicular to a drive shaft of the
compressor
[0082] FIG. 44 is a perspective view of the view in FIG. 43 of the
compressor in FIG. 39.
[0083] FIG. 45 is cross-sectional view of a discharge manifold of
the compressor in FIG. 39.
[0084] FIG. 46 is perspective view of the discharge manifold in
FIG. 45.
[0085] FIG. 47 is an end view of the discharge manifold in FIG.
45.
[0086] FIG. 48 is partial, cross-sectional, perspective view of the
compressor in FIG. 39, showing the hydrostatic bearing
arrangement.
[0087] FIG. 49 is perspective view of the hydrostatic bearings and
gate of the compressor in FIG. 39.
[0088] FIG. 50 is diagrammatic view of the hydrostatic bearing
arrangement of the compressor in FIG. 39.
[0089] FIG. 51 is a resistance flow diagram of the hydrostatic
bearings of the compressor in FIG. 39.
[0090] FIG. 52 is a partial cross-sectional view of FIG. 40.
[0091] FIG. 53 is a partial cross-sectional view of a compressor
according to an alternative embodiment.
[0092] FIG. 54 is an enlarged, partial, cross-sectional view of
FIG. 52.
[0093] FIG. 55 is a perspective view of a compressor according to
an alternative embodiment, with a cam casing removed to display
internal components.
[0094] FIG. 56 is a cross-sectional view of the compressor in FIG.
55, taken in a plane that is perpendicular to a drive shaft of the
compressor.
[0095] FIG. 57 is a cross-sectional view of the compressor in FIG.
55, taken along an axis of the compressor's drive shaft.
[0096] FIG. 58 is a perspective view of the compressor in FIG. 55,
showing a cam casing.
[0097] FIG. 59 is a perspective view of a compressor according to
an alternative embodiment.
[0098] FIG. 60 is a cross-sectional view of the compressor in FIG.
59, taken along an axis of the compressor's drive shaft.
[0099] FIGS. 61 and 62 are cross-sectional views of a compressor
according to an alternative embodiment, with the cross-sections
taken perpendicular to an axis of a drive shaft of the
compressor.
[0100] FIGS. 63-65 are end views of the compressor of FIGS. 61 and
62, taken at different points in the compression cycle.
[0101] FIG. 66 is a cross-sectional view of a compressor according
to an alternative embodiment, taken along an axis of the
compressor's drive shaft.
[0102] FIG. 67 is a cross-sectional end view of the rotor of the
compressor in FIG. 39, with the cross-section taken perpendicular
to the drive shaft.
[0103] FIG. 68 is a cross-sectional view of the rotor and drive
shaft in FIG. 67, with the cross-section taken along the line 68-68
in FIG. 67.
[0104] FIG. 69 is a partial cross-sectional view of a compressor
according to an alternative embodiment, with the cross-section
taken along an axis of the compressor's drive shaft.
[0105] FIG. 70 is a side view of a compressor according to an
alternative embodiment;
[0106] FIG. 71 is an end view of the compressor in FIG. 70;
[0107] FIG. 72 is a perspective side view of the compressor in FIG.
70;
[0108] FIG. 73 is a cross-sectional view of the compressor in FIG.
70, taken along the line 73-73 in FIG. 70; and
[0109] FIG. 74 is a partial, magnified cross-sectional view of FIG.
73.
DETAILED DESCRIPTION OF THE EMBODIMENTS
[0110] To the extent that the following terms are utilized herein,
the following definitions are applicable:
[0111] Balanced rotation: the center of mass of the rotating mass
is located on the axis of rotation.
[0112] Chamber volume: any volume that can contain fluids for
compression.
[0113] Compressor: a device used to increase the pressure of a
compressible fluid. The fluid can be either gas or vapor, and can
have a wide molecular weight range.
[0114] Concentric: the center or axis of one object coincides with
the center or axis of a second object
[0115] Concentric rotation: rotation in which one object's center
of rotation is located on the same axis as the second object's
center of rotation.
[0116] Positive displacement compressor: a compressor that collects
a fixed volume of gas within a chamber and compresses it by
reducing the chamber volume.
[0117] Proximate: sufficiently close to restrict fluid flow between
high pressure and low pressure regions. Restriction does not need
to be absolute; some leakage is acceptable.
[0118] Rotor: A rotating element driven by a mechanical force to
rotate about an axis. As used in a compressor design, the rotor
imparts energy to a fluid.
[0119] Rotary compressor: A positive-displacement compressor that
imparts energy to the gas being compressed by way of an input shaft
moving a single or multiple rotating elements
[0120] FIGS. 1 through 7 show external views of an embodiment of
the present invention in which a rotary compressor includes spring
backed cam drive gate positioning system. Main housing 100 includes
a main casing 110 and end plates 120, each of which includes a hole
through which drive shaft 140 passes axially. Liquid injector
assemblies 130 are located on holes in the main casing 110. The
main casing includes a hole for the inlet flange 160, and a hole
for the gate casing 150.
[0121] Gate casing 150 is connected to and positioned below main
casing 110 at a hole in main casing 110. The gate casing 150 is
comprised of two portions: an inlet side 152 and an outlet side
154. Other embodiments of gate casing 150 may only consist of a
single portion. As shown in FIG. 28, the outlet side 154 includes
outlet ports 435, which are holes which lead to outlet valves 440.
Alternatively, an outlet valve assembly may be used.
[0122] Referring back to FIGS. 1-7, the spring-backed cam drive
gate positioning system 200 is attached to the gate casing 150 and
drive shaft 140. The gate positioning system 200 moves gate 600 in
conjunction with the rotation of rotor 500. A movable assembly
includes gate struts 210 and cam struts 230 connected to gate
support arm 220 and bearing support plate 156. The bearing support
plate 156 seals the gate casing 150 by interfacing with the inlet
and outlet sides through a bolted gasket connection. Bearing
support plate 156 is shaped to seal gate casing 150, mount bearing
housings 270 in a sufficiently parallel manner, and constrain
compressive springs 280. In one embodiment, the interior of the
gate casing 150 is hermetically sealed by the bearing support plate
156 with o-rings, gaskets, or other sealing materials. Other
embodiments may support the bearings at other locations, in which
case an alternate plate may be used to seal the interior of the
gate casing. Shaft seals, mechanical seals, or other sealing
mechanisms may be used to seal around the gate struts 210 which
penetrate the bearing support plate 156 or other sealing plate.
Bearing housings 270, also known as pillow blocks, are concentric
to the gate struts 210 and the cam struts 230.
[0123] In the illustrated embodiment, the compressing structure
comprises a rotor 500. However, according to alternative
embodiments, alternative types of compressing structures (e.g.,
gears, screws, pistons, etc.) may be used in connection with the
compression chamber to provide alternative compressors according to
alternative embodiments of the invention.
[0124] Two cam followers 250 are located tangentially to each cam
240, providing a downward force on the gate. Drive shaft 140 turns
cams 240, which transmits force to the cam followers 250. The cam
followers 250 may be mounted on a through shaft, which is supported
on both ends, or cantilevered and only supported on one end. The
cam followers 250 are attached to cam follower supports 260, which
transfer the force into the cam struts 230. As cams 240 turn, the
cam followers 250 are pushed down, thus moving the cam struts 230
down. This moves the gate support arm 220 and the gate strut 210
down. This, in turn, moves the gate 600 down.
[0125] Springs 280 provide a restorative upward force to keep the
gate 600 timed appropriately to seal against the rotor 500. As the
cams 240 continue to turn and no longer effectuate a downward force
on the cam followers 250, springs 280 provide an upward force. As
shown in this embodiment, compression springs are utilized. As one
of ordinary skill in the art would appreciate, tension springs and
the shape of the bearing support plate 156 may be altered to
provide for the desired upward or downward force. The upward force
of the springs 280 pushes the cam follower support 260 and thus the
gate support arm 220 up which in turn moves the gate 600 up.
[0126] Due to the varying pressure angle between the cam followers
250 and cams 240, the preferred embodiment may utilize an exterior
cam profile that differs from the rotor 500 profile. This variation
in profile allows for compensation for the changing pressure angle
to ensure that the tip of the gate 600 remains proximate to the
rotor 500 throughout the entire compression cycle.
[0127] Line A in FIGS. 3, 6, and 7 shows the location for the
cross-sectional view of the compressor in FIG. 8. As shown in FIG.
8, the main casing 110 has a cylindrical shape. Liquid injector
housings 132 are attached to, or may be cast as a part of, the main
casing 110 to provide for openings in the rotor casing 400. Because
it is cylindrically shaped in this embodiment, the rotor casing 400
may also be referenced as the cylinder. The interior wall defines a
rotor casing volume 410 (also referred to as the compression
chamber). The rotor 500 concentrically rotates with drive shaft 140
and is affixed to the drive shaft 140 by way of key 540 and press
fit. Alternate methods for affixing the rotor 500 to the drive
shaft 140, such as polygons, splines, or a tapered shaft may also
be used.
[0128] FIG. 9 shows an embodiment of the present invention in which
a timing belt with spring gate positioning system is utilized. This
embodiment 290 incorporates two timing belts 292 each of which is
attached to the drive shaft 140 by way of sheaves 294. The timing
belts 292 are attached to secondary shafts 142 by way of sheaves
295. Gate strut springs 296 are mounted around gate struts. Rocker
arms 297 are mounted to rocker arm supports 299. The sheaves 295
are connected to rocker arm cams 293 to push the rocker arms 297
down. As the inner rings push down on one side of the rocker arms
297, the other side pushes up against the gate support bar 298. The
gate support bar 298 pushes up against the gate struts and gate
strut springs 296. This moves the gate up. The springs 296 provide
a downward force pushing the gate down.
[0129] FIGS. 10 through 17 show external views of a rotary
compressor embodiment utilizing a dual cam follower gate
positioning system. The main housing 100 includes a main casing 110
and end plates 120, each of which includes a hole through which a
drive shaft 140 passes axially. Liquid injector assemblies 130 are
located on holes in the main casing 110. The main casing 110 also
includes a hole for the inlet flange 160 and a hole for the gate
casing 150. The gate casing 150 is mounted to and positioned below
the main casing 110 as discussed above.
[0130] A dual cam follower gate positioning system 300 is attached
to the gate casing 150 and drive shaft 140. The dual cam follower
gate positioning system 300 moves the gate 600 in conjunction with
the rotation of the rotor 500. In a preferred embodiment, the size
and shape of the cams is nearly identical to the rotor in
cross-sectional size and shape. In other embodiments, the rotor,
cam shape, curvature, cam thickness, and variations in the
thickness of the lip of the cam may be adjusted to account for
variations in the attack angle of the cam follower. Further, large
or smaller cam sizes may be used. For example, a similar shape but
smaller size cam may be used to reduce roller speeds.
[0131] A movable assembly includes gate struts 210 and cam struts
230 connected to gate support arm 220 and bearing support plate
156. In this embodiment, the bearing support plate 157 is straight.
As one of ordinary skill in the art would appreciate, the bearing
support plate can utilize different geometries, including
structures designed to or not to perform sealing of the gate casing
150. In this embodiment, the bearing support plate 157 serves to
seal the bottom of the gate casing 150 through a bolted gasket
connection. Bearing housings 270, also known as pillow blocks, are
mounted to bearing support plate 157 and are concentric to the gate
struts 210 and the cam struts 230. In certain embodiments, the
components comprising this movable assembly may be optimized to
reduce weight, thereby reducing the force necessary to achieve the
necessary acceleration to keep the tip of gate 600 proximate to the
rotor 500. Weight reduction could additionally and/or alternatively
be achieved by removing material from the exterior of any of the
moving components, as well as by hollowing out moving components,
such as the gate struts 210 or the gate 600.
[0132] Drive shaft 140 turns cams 240, which transmit force to the
cam followers 250, including upper cam followers 252 and lower cam
followers 254. The cam followers 250 may be mounted on a through
shaft, which is supported on both ends, or cantilevered and only
supported on one end. In this embodiment, four cam followers 250
are used for each cam 240. Two lower cam followers 252 are located
below and follow the outside edge of the cam 240. They are mounted
using a through shaft. Two upper cam followers 254 are located
above the previous two and follow the inside edge of the cams 240.
They are mounted using a cantilevered connection.
[0133] The cam followers 250 are attached to cam follower supports
260, which transfer the force into the cam struts 230. As the cams
240 turn, the cam struts 230 move up and down. This moves the gate
support arm 220 and gate struts 210 up and down, which in turn,
moves the gate 600 up and down.
[0134] Line A in FIGS. 11, 12, 15, and 16 show the location for the
cross-sectional view of the compressor in FIG. 17. As shown in FIG.
17, the main casing 110 has a cylindrical shape. Liquid injector
housings 132 are attached to or may be cast as a part of the main
casing 110 to provide for openings in the rotor casing 400. The
rotor 500 concentrically rotates around drive shaft 140.
[0135] An embodiment using a belt driven system 310 is shown in
FIG. 18. Timing belts 292 are connected to the drive shaft 140 by
way of sheaves 294. The timing belts 292 are each also connected to
secondary shafts 142 by way of another set of sheaves 295. The
secondary shafts 142 drive the external cams 240, which are placed
below the gate casing 150 in this embodiment. Sets of upper and
lower cam followers 254 and 252 are applied to the cams 240, which
provide force to the movable assembly including gate struts 210 and
gate support arm 220. As one of ordinary skill in the art would
appreciate, belts may be replaced by chains or other materials.
[0136] An embodiment of the present invention using an offset gate
guide system is shown in FIGS. 19 through 22 and 33. Outlet of the
compressed gas and injected fluid is achieved through a ported gate
system 602 comprised of two parts bolted together to allow for
internal lightening features. Fluid passes through channels 630 in
the upper portion of the gate 602 and travels to the lengthwise
sides to outlet through an exhaust port 344 in a timed manner with
relation to the angle of rotation of the rotor 500 during the
cycle. Discrete point spring-backed scraper seals 326 provide
sealing of the gate 602 in the single piece gate casing 336. Liquid
injection is achieved through a variety of flat spray nozzles 322
and injector nozzles 130 across a variety of liquid injector port
324 locations and angles.
[0137] Reciprocating motion of the two-piece gate 602 is controlled
through the use of an offset spring-backed cam follower control
system 320 to achieve gate motion in concert with rotor rotation.
Single cams 342 drive the gate system downwards through the
transmission of force on the cam followers 250 through the cam
struts 338. This results in controlled motion of the crossarm 334,
which is connected by bolts (some of which are labeled as 328) with
the two-piece gate 602. The crossarm 334 mounted linear bushings
330, which reciprocate along the length of cam shafts 332, control
the motion of the gate 602 and the crossarm 334. The cam shafts 332
are fixed in a precise manner to the main casing through the use of
cam shaft support blocks 340. Compression springs 346 are utilized
to provide a returning force on the crossarm 334, allowing the cam
followers 250 to maintain constant rolling contact with the cams,
thereby achieving controlled reciprocating motion of the two-piece
gate 602.
[0138] FIG. 23 shows an embodiment using a linear actuator system
350 for gate positioning. A pair of linear actuators 352 is used to
drive the gate. In this embodiment, it is not necessary to
mechanically link the drive shaft to the gate as with other
embodiments. The linear actuators 352 are controlled so as to raise
and lower the gate in accordance with the rotation of the rotor.
The actuators may be electronic, hydraulic, belt-driven,
electromagnetic, gas-driven, variable-friction, or other means. The
actuators may be computer controlled or controlled by other
means.
[0139] FIGS. 24A and B show a magnetic drive system 360. The gate
system may be driven, or controlled, in a reciprocating motion
through the placement of magnetic field generators, whether they
are permanent magnets or electromagnets, on any combination of the
rotor 500, gate 600, and/or gate casing 150. The purpose of this
system is to maintain a constant distance from the tip of the gate
600 to the surface of the rotor 500 at all angles throughout the
cycle. In a preferred magnetic system embodiment, permanent magnets
366 are mounted into the ends of the rotor 500 and retained. In
addition, permanent magnets 364 are installed and retained in the
gate 600. Poles of the magnets are aligned so that the magnetic
force generated between the rotor's magnets 366 and the gate's
magnets 364 is a repulsive force, forcing the gate 600 down
throughout the cycle to control its motion and maintain constant
distance. To provide an upward, returning force on the gate 600,
additional magnets (not shown) are installed into the bottom of the
gate 600 and the bottom of the gate casing 150 to provide an
additional repulsive force. The magnetic drive systems are balanced
to precisely control the gate's reciprocating motion.
[0140] Alternative embodiments may use an alternate pole
orientation to provide attractive forces between the gate and rotor
on the top portion of the gate and attractive forces between the
gate and gate casing on the bottom portion of the gate. In place of
the lower magnet system, springs may be used to provide a repulsive
force. In each embodiment, electromagnets may be used in place of
permanent magnets. In addition, switched reluctance electromagnets
may also be utilized. In another embodiment, electromagnets may be
used only in the rotor and gate. Their poles may switch at each
inflection point of the gate's travel during its reciprocating
cycle, allowing them to be used in an attractive and repulsive
method.
[0141] Alternatively, direct hydraulic or indirect hydraulic
(hydropneumatic) can be used to apply motive force/energy to the
gate to drive it and position it adequately. Solenoid or other flow
control valves can be used to feed and regulate the position and
movement of the hydraulic or hydropneumatic elements. Hydraulic
force may be converted to mechanical force acting on the gate
through the use of a cylinder based or direct hydraulic actuators
using membranes/diaphragms.
[0142] FIG. 25 shows an embodiment using a scotch yoke gate
positioning system 370. Here, a pair of scotch yokes 372 is
connected to the drive shaft and the bearing support plate. A
roller rotates at a fixed radius with respect to the shaft. The
roller follows a slot within the yoke 372, which is constrained to
a reciprocating motion. The yoke geometry can be manipulated to a
specific shape that will result in desired gate dynamics.
[0143] As one of skill in the art would appreciate, these
alternative drive mechanisms do not require any particular number
of linkages between the drive shaft and the gate. For example, a
single spring, belt, linkage bar, or yoke could be used. Depending
on the design implementation, more than two such elements could be
used.
[0144] FIGS. 26A-26F show a compression cycle of an embodiment
utilizing a tip seal 620. As the drive shaft 140 turns, the rotor
500 and gate strut 210 push up gate 600 so that it is timed with
the rotor 500. As the rotor 500 turns clockwise, the gate 600 rises
up until the rotor 500 is in the 12 o'clock position shown in FIG.
26C. As the rotor 500 continues to turn, the gate 600 moves
downward until it is back at the 6 o'clock position in FIG. 26F.
The gate 600 separates the portion of the cylinder that is not
taken up by rotor 500 into two components: an intake component 412
and a compression component 414. In one embodiment, tip seal 620
may not be centered within the gate 600, but may instead be shifted
towards one side so as to minimize the area on the top of the gate
on which pressure may exert a downwards force on the gate. This may
also have the effect of minimizing the clearance volume of the
system. In another embodiment, the end of the tip seal 620
proximate to the rotor 500 may be rounded, so as to accommodate the
varying contact angle that will be encountered as the tip seal 620
contacts the rotor 500 at different points in its rotation.
[0145] FIGS. 26A-F depict steady state operation. Accordingly, in
FIG. 26A, where the rotor 500 is in the 6 o'clock position, the
compression volume 414, which constitutes a subset of the rotor
casing volume 410, already has received fluid. In FIG. 26B, the
rotor 500 has turned clockwise and gate 600 has risen so that the
tip seal 620 makes contact with the rotor 500 to separate the
intake volume 412, which also constitutes a subset of the rotor
casing volume 410, from the compression volume 414. Embodiments
using the roller tip 650 discussed below instead of tip seal 620
would operate similarly. As the rotor 500 turns, as shown further
in FIGS. 26C-E, the intake volume 412 increases, thereby drawing in
more fluid from inlet 420, while the compression volume 414
decreases. As the volume of the compression volume 414 decreases,
the pressure increases. The pressurized fluid is then expelled by
way of an outlet 430. At a point in the compression cycle when a
desired high pressure is reached, the outlet valve opens and the
high pressure fluid can leave the compression volume 414. In this
embodiment, the valve outputs both the compressed gas and the
liquid injected into the compression chamber.
[0146] FIGS. 27A-27F show an embodiment in which the gate 600 does
not use a tip seal. Instead, the gate 600 is timed to be proximate
to the rotor 500 as it turns. The close proximity of the gate 600
to the rotor 500 leaves only a very small path for high pressure
fluid to escape. Close proximity in conjunction with the presence
of liquid (due to the liquid injectors 136 or an injector placed in
the gate itself) allow the gate 600 to effectively create an intake
fluid component 412 and a compression component 414. Embodiments
incorporating notches 640 would operate similarly.
[0147] FIG. 28 shows a cross-sectional perspective view of the
rotor casing 400, the rotor 500, and the gate 600. The inlet port
420 shows the path that gas can enter. The outlet 430 is comprised
of several holes that serve as outlet ports 435 that lead to outlet
valves 440. The gate casing 150 consists of an inlet side 152 and
an outlet side 154. A return pressure path (not shown) may be
connected to the inlet side 152 of the gate casing 150 and the
inlet port 420 to ensure that there is no back pressure build up
against gate 600 due to leakage through the gate seals. As one of
ordinary skill in the art would appreciate, it is desirable to
achieve a hermetic seal, although perfect hermetic sealing is not
necessary.
[0148] In alternate embodiments, the outlet ports 435 may be
located in the rotor casing 400 instead of the gate casing 150.
They may be located at a variety of different locations within the
rotor casing. The outlet valves 440 may be located closer to the
compression chamber, effectively minimizing the volume of the
outlet ports 430, to minimize the clearance volume related to these
outlet ports. A valve cartridge may be used which houses one or
more outlet valves 440 and connects directly to the rotor casing
400 or gate casing 150 to align the outlet valves 440 with outlet
ports 435. This may allow for ease of installing and removing the
outlet valves 440.
[0149] FIG. 29 shows an alternative embodiment in which flat spray
liquid injector housings 170 are located on the main casing 110 at
approximately the 3 o'clock position. These injectors can be used
to inject liquid directly onto the inlet side of the gate 600,
ensuring that it does not reach high temperatures. These injectors
also help to provide a coating of liquid on the rotor 500, helping
to seal the compressor.
[0150] As discussed above, the preferred embodiments utilize a
rotor that concentrically rotates within a rotor casing. In the
preferred embodiment, the rotor 500 is a right cylinder with a
non-circular cross-section that runs the length of the main casing
110. FIG. 30 shows a cross-sectional view of the sealing and
non-sealing portions of the rotor 500. The profile of the rotor 500
is comprised of three sections. The radii in sections I and III are
defined by a cycloidal curve. This curve also represents the rise
and fall of the gate and defines an optimum acceleration profile
for the gate. Other embodiments may use different curve functions
to define the radius such as a double harmonic function. Section II
employs a constant radius 570, which corresponds to the maximum
radius of the rotor. The minimum radius 580 is located at the
intersection of sections I and III, at the bottom of rotor 500. In
a preferred embodiment, .PHI. is 23.8 degrees. In alternative
embodiments, other angles may be utilized depending on the desired
size of the compressor, the desired acceleration of the gate, and
desired sealing area.
[0151] The radii of the rotor 500 in one preferred embodiment can
be calculated using the following functions:
r .function. ( t ) = { r I = r min .times. h .function. [ t I T sin
.function. ( 2 .times. .times. .times. t I T ) ] r II = r max r III
= r min + h .function. [ t III T + sin .function. ( 2 .times.
.times. .times. t III T ) ] ##EQU00001##
[0152] According to an alternative embodiment, the radii of the
rotor 500 is calculated as a 3-4-5-polynomial function.
[0153] In a preferred embodiment, the rotor 500 is symmetrical
along one axis. It may generally resemble a cross-sectional egg
shape. The rotor 500 includes a hole 530 in which the drive shaft
140 and a key 540 may be mounted. The rotor 500 has a sealing
section 510, which is the outer surface of the rotor 500
corresponding to section II, and a non-sealing section 520, which
is the outer surface of the rotor 500 corresponding to sections I
and III. The sections I and III have a smaller radius than sections
II creating a compression volume. The sealing portion 510 is shaped
to correspond to the curvature of the rotor casing 400, thereby
creating a dwell seal that effectively minimizes communication
between the outlet 430 and inlet 420. Physical contact is not
required for the dwell seal. Instead, it is sufficient to create a
tortuous path that minimizes the amount of fluid that can pass
through. In a preferred embodiment, the gap between the rotor and
the casing in this embodiment is less than 0.008 inches. As one of
ordinary skill in the art would appreciate, this gap may be altered
depending on tolerances, both in machining the rotor 500 and rotor
housing 400, temperature, material properties, and other specific
application requirements.
[0154] Additionally, as discussed below, liquid is injected into
the compression chamber. By becoming entrained in the gap between
the sealing portion 510 and the rotor casing 400, the liquid can
increase the effectiveness of the dwell seal.
[0155] As shown in FIG. 31A, the rotor 500 is balanced with cut out
shapes and counterweights. Holes, some of which are marked as 550,
lighten the rotor 500. These lightening holes may be filled with a
low density material to ensure that liquid cannot encroach into the
rotor interior. Alternatively, caps may be placed on the ends of
rotor 500 to seal the lightening holes. Counterweights, one of
which is labeled as 560, are made of a denser material than the
remainder of the rotor 500. The shapes of the counterweights can
vary and do not need to be cylindrical.
[0156] The rotor design provides several advantages. As shown in
the embodiment of FIG. 31A, the rotor 500 includes 7 cutout holes
550 on one side and two counterweights 560 on the other side to
allow the center of mass to match the center of rotation. An
opening 530 includes space for the drive shaft and a key. This
weight distribution is designed to achieve balanced, concentric
motion. The number and location of cutouts and counterweights may
be changed depending on structural integrity, weight distribution,
and balanced rotation parameters. In various embodiments, cutouts
and/or counterweights or neither may be used required to achieve
balanced rotor rotation.
[0157] The cross-sectional shape of the rotor 500 allows for
concentric rotation about the drive shaft's axis of rotation, a
dwell seal 510 portion, and open space on the non-sealing side for
increased gas volume for compression. Concentric rotation provides
for rotation about the drive shaft's principal axis of rotation and
thus smoother motion and reduced noise.
[0158] An alternative rotor design 502 is shown in FIG. 31B. In
this embodiment, a different arc of curvature is implemented
utilizing three holes 550 and a circular opening 530. Another
alternative design 504 is shown in FIG. 31C. Here, a solid rotor
shape is used and a larger hole 530 (for a larger drive shaft) is
implemented. Yet another alternative rotor design 506 is shown in
FIG. 31D incorporating an asymmetrical shape, which would smooth
the volume reduction curve, allowing for increased time for heat
transfer to occur at higher pressures. Alternative rotor shapes may
be implemented for different curvatures or needs for increased
volume in the compression chamber.
[0159] The rotor surface may be smooth in embodiments with
contacting tip seals to minimize wear on the tip seal. In
alternative embodiments, it may be advantageous to put surface
texture on the rotor to create turbulence that may improve the
performance of non-contacting seals. In other embodiments, the
rotor casing's interior cylindrical wall may further be textured to
produce additional turbulence, both for sealing and heat transfer
benefits. This texturing could be achieved through machining of the
parts or by utilizing a surface coating. Another method of
achieving the texture would be through blasting with a waterjet,
sandblast, or similar device to create an irregular surface.
[0160] The main casing 110 may further utilize a removable cylinder
liner. This liner may feature microsurfacing to induce turbulence
for the benefits noted above. The liner may also act as a wear
surface to increase the reliability of the rotor and casing. The
removable liner could be replaced at regular intervals as part of a
recommended maintenance schedule. The rotor may also include a
liner. Sacrifical or wear-in coatings may be used on the rotor 500
or rotor casing 400 to correct for manufacturing defects in
ensuring the preferred gap is maintained along the sealing portion
510 of the rotor 500.
[0161] The exterior of the main casing 110 may also be modified to
meet application specific parameters. For example, in subsea
applications, the casing may require to be significantly thickened
to withstand exterior pressure, or placed within a secondary
pressure vessel. Other applications may benefit from the exterior
of the casing having a rectangular or square profile to facilitate
mounting exterior objects or stacking multiple compressors. Liquid
may be circulated in the casing interior to achieve additional heat
transfer or to equalize pressure in the case of subsea applications
for example.
[0162] As shown in FIG. 32A and B, the combination of the rotor 500
(here depicted with rotor end caps 590), the gate 600, and drive
shaft 140, provide for a more efficient manner of compressing
fluids in a cylinder. The gate is aligned along the length of the
rotor to separate and define the inlet portion and compression
portion as the rotor turns.
[0163] The drive shaft 140 is mounted to endplates 120 in the
preferred embodiment using one spherical roller bearing in each
endplate 120. More than one bearing may be used in each endplate
120, in order to increase total load capacity. A grease pump (not
shown) is used to provide lubrication to the bearings. Various
types of other bearings may be utilized depending on application
specific parameters, including roller bearings, ball bearings,
needle bearings, conical bearings, cylindrical bearings, journal
bearings, etc. Different lubrication systems using grease, oil, or
other lubricants may also be used. Further, dry lubrication systems
or materials may be used. Additionally, applications in which
dynamic imbalance may occur may benefit from multi-bearing
arrangements to support stray axial loads.
[0164] Operation of gates in accordance with embodiments of the
present invention are shown in FIGS. 8, 17, 22, 24B, 26A-F, 27A-F,
28, 32A-B, and 33-36. As shown in FIGS. 26A-F and 27A-F, gate 600
creates a pressure boundary between an intake volume 412 and a
compression volume 414. The intake volume 412 is in communication
with the inlet 420. The compression volume 414 is in communication
with the outlet 430. Resembling a reciprocating, rectangular
piston, the gate 600 rises and falls in time with the turning of
the rotor 500.
[0165] The gate 600 may include an optional tip seal 620 that makes
contact with the rotor 500, providing an interface between the
rotor 500 and the gate 600. Tip seal 620 consists of a strip of
material at the tip of the gate 600 that rides against rotor 500.
The tip seal 620 could be made of different materials, including
polymers, graphite, and metal, and could take a variety of
geometries, such as a curved, flat, or angled surface. The tip seal
620 may be backed by pressurized fluid or a spring force provided
by springs or elastomers. This provides a return force to keep the
tip seal 620 in sealing contact with the rotor 500.
[0166] Different types of contacting tips may be used with the gate
600. As shown in FIG. 35, a roller tip 650 may be used. The roller
tip 650 rotates as it makes contact with the turning rotor 500.
Also, tips of differing strengths may be used. For example, a tip
seal 620 or roller tip 650 may be made of softer metal that would
gradually wear down before the rotor 500 surfaces would wear.
[0167] Alternatively, a non-contacting seal may be used.
Accordingly, the tip seal may be omitted. In these embodiments, the
topmost portion of the gate 600 is placed proximate, but not
necessarily in contact with, the rotor 500 as it turns. The amount
of allowable gap may be adjusted depending on application
parameters.
[0168] As shown in FIGS. 34A and 34B, in an embodiment in which the
tip of the gate 600 does not contact the rotor 500, the tip may
include notches 640 that serve to keep gas pocketed against the tip
of the gate 600. The entrained fluid, in either gas or liquid form,
assists in providing a non-contacting seal. As one of ordinary
skill in the art would appreciate, the number and size of the
notches is a matter of design choice dependent on the compressor
specifications.
[0169] Alternatively, liquid may be injected from the gate itself.
As shown in FIG. 36, a cross-sectional view of a portion of a gate,
one or more channels 660 from which a fluid may pass may be built
into the gate. In one such embodiment, a liquid can pass through a
plurality of channels 660 to form a liquid seal between the topmost
portion of the gate 600 and the rotor 500 as it turns. In another
embodiment, residual compressed fluid may be inserted through one
or more channels 660. Further still, the gate 600 may be shaped to
match the curvature of portions of the rotor 500 to minimize the
gap between the gate 600 and the rotor 500.
[0170] Preferred embodiments enclose the gate in a gate casing. As
shown in FIGS. 8 and 17, the gate 600 is encompassed by the gate
casing 150, including notches, one of which is shown as item 158.
The notches hold the gate seals, which ensure that the compressed
fluid will not release from the compression volume 414 through the
interface between gate 600 and gate casing 150 as gate 600 moves up
and down. The gate seals may be made of various materials,
including polymers, graphite or metal. A variety of different
geometries may be used for these seals. Various embodiments could
utilize different notch geometries, including ones in which the
notches may pass through the gate casing, in part or in full.
[0171] In alternate embodiments, the seals could be placed on the
gate 600 instead of within the gate casing 150. The seals would
form a ring around the gate 600 and move with the gate relative to
the casing 150, maintaining a seal against the interior of the gate
casing 150. The location of the seals may be chosen such that the
center of pressure on the gate 600 is located on the portion of the
gate 600 inside of the gate casing 150, thus reducing or
eliminating the effect of a cantilevered force on the portion of
the gate 600 extending into the rotor casing 400. This may help
eliminate a line contact between the gate 600 and gate casing 150
and instead provide a surface contact, allowing for reduced
friction and wear. One or more wear plates may be used on the gate
600 to contact the gate casing 150. The location of the seals and
wear plates may be optimized to ensure proper distribution of
forces across the wear plates.
[0172] The seals may use energizing forces provided by springs or
elastomers with the assembly of the gate casing 150 inducing
compression on the seals. Pressurized fluid may also be used to
energize the seals.
[0173] The gate 600 is shown with gate struts 210 connected to the
end of the gate. In various embodiments, the gate 600 may be
hollowed out such that the gate struts 210 can connect to the gate
600 closer to its tip. This may reduce the amount of thermal
expansion encountered in the gate 600. A hollow gate also reduces
the weight of the moving assembly and allows oil or other
lubricants and coolants to be splashed into the interior of the
gate to maintain a cooler temperature. The relative location of
where the gate struts 210 connect to the gate 600 and where the
gate seals are located may be optimized such that the deflection
modes of the gate 600 and gate struts 210 are equal, allowing the
gate 600 to remain parallel to the interior wall of the gate casing
150 when it deflects due to pressure, as opposed to rotating from
the pressure force. Remaining parallel may help to distribute the
load between the gate 600 and gate casing 150 to reduce friction
and wear.
[0174] A rotor face seal may also be placed on the rotor 500 to
provide for an interface between the rotor 500 and the endplates
120. An outer rotor face seal is placed along the exterior edge of
the rotor 500, preventing fluid from escaping past the end of the
rotor 500. A secondary inner rotor face seal is placed on the rotor
face at a smaller radius to prevent any fluid that escapes past the
outer rotor face seal from escaping the compressor entirely. This
seal may use the same or other materials as the gate seal. Various
geometries may be used to optimize the effectiveness of the seals.
These seals may use energizing forces provided by springs,
elastomers or pressurized fluid. Lubrication may be provided to
these rotor face seals by injecting oil or other lubricant through
ports in the endplates 120.
[0175] Along with the seals discussed herein, the surfaces those
seals contact, known as counter-surfaces, may also be considered.
In various embodiments, the surface finish of the counter-surface
may be sufficiently smooth to minimize friction and wear between
the surfaces. In other embodiments, the surface finish may be
roughened or given a pattern such as cross-hatching to promote
retention of lubricant or turbulence of leaking fluids. The
counter-surface may be composed of a harder material than the seal
to ensure the seal wears faster than the counter-surface, or the
seal may be composed of a harder material than the counter-surface
to ensure the counter-surface wears faster than the seal. The
desired physical properties of the counter-surface (surface
roughness, hardness, etc.) may be achieved through material
selection, material finishing techniques such as quenching,
tempering, or work hardening, or selection and application of
coatings that achieve the desired characteristics. Final
manufacturing processes, such as surface grinding, may be performed
before or after coatings are applied. In various embodiments, the
counter-surface material may be steel or stainless steel. The
material may be hardened via quenching or tempering. A coating may
be applied, which could be chrome, titanium nitride, silicon
carbide, or other materials.
[0176] Minimizing the possibility of fluids leaking to the exterior
of the main housing 100 is desirable. Various seals, such as
gaskets and o-rings, are used to seal external connections between
parts. For example, in a preferred embodiment, a double o-ring seal
is used between the main casing 110 and endplates 120. Further
seals are utilized around the drive shaft 140 to prevent leakage of
any fluids making it past the rotor face seals. A lip seal is used
to seal the drive shaft 140 where it passes through the endplates
120. In various embodiments, multiple seals may be used along the
drive shaft 140 with small gaps between them to locate vent lines
and hydraulic packings to reduce or eliminate gas leakage exterior
to the compression chamber. Other forms of seals could also be
used, such as mechanical or labyrinth seals.
[0177] It is desirable to achieve near isothermal compression. To
provide cooling during the compression process, liquid injection is
used. In preferred embodiments, the liquid is atomized to provide
increased surface area for heat absorption. In other embodiments,
different spray applications or other means of injecting liquids
may be used.
[0178] Liquid injection is used to cool the fluid as it is
compressed, increasing the efficiency of the compression process.
Cooling allows most of the input energy to be used for compression
rather than heat generation in the gas. The liquid has dramatically
superior heat absorption characteristics compared to gas, allowing
the liquid to absorb heat and minimize temperature increase of the
working fluid, achieving near isothermal compression. As shown in
FIGS. 8 and 17, liquid injector assemblies 130 are attached to the
main casing 110. Liquid injector housings 132 include an adapter
for the liquid source 134 (if it is not included with the nozzle)
and a nozzle 136. Liquid is injected by way of a nozzle 136
directly into the rotor casing volume 410.
[0179] The amount and timing of liquid injection may be controlled
by a variety of implements including a computer-based controller
capable of measuring the liquid drainage rate, liquid levels in the
chamber, and/or any rotational resistance due to liquid
accumulation through a variety of sensors. Valves or solenoids may
be used in conjunction with the nozzles to selectively control
injection timing. Variable orifice control may also be used to
regulate the amount of liquid injection and other
characteristics.
[0180] Analytical and experimental results are used to optimize the
number, location, and spray direction of the injectors 136. These
injectors 136 may be located in the periphery of the cylinder.
Liquid injection may also occur through the rotor or gate. The
current embodiment of the design has two nozzles located at 12
o'clock and 10 o'clock. Different application parameters will also
influence preferred nozzle arrays.
[0181] Because the heat capacity of liquids is typically much
higher than gases, the heat is primarily absorbed by the liquid,
keeping gas temperatures lower than they would be in the absence of
such liquid injection.
[0182] When a fluid is compressed, the pressure times the volume
raised to a polytropic exponent remains constant throughout the
cycle, as seen in the following equation:
P*V.sup.n=Constant
[0183] In polytropic compression, two special cases represent the
opposing sides of the compression spectrum. On the high end,
adiabatic compression is defined by a polytropic constant of n=1.4
for air, or n=1.28 for methane. Adiabatic compression is
characterized by the complete absence of cooling of the working
fluid (isentropic compression is a subset of adiabatic compression
in which the process is reversible). This means that as the volume
of the fluid is reduced, the pressure and temperature each rise
accordingly. It is an inefficient process due to the exorbitant
amount of energy wasted in the generation of heat in the fluid,
which often needs to be cooled down again later. Despite being an
inefficient process, most conventional compression technology,
including reciprocating piston and centrifugal type compressors are
essentially adiabatic. The other special case is isothermal
compression, where n=1. It is an ideal compression cycle in which
all heat generated in the fluid is transmitted to the environment,
maintaining a constant temperature in the working fluid. Although
it represents an unachievable perfect case, isothermal compression
is useful in that it provides a lower limit to the amount of energy
required to compress a fluid.
[0184] FIG. 37 shows a sample pressure-volume (P-V) curve comparing
several different compression processes. The isothermal curve shows
the theoretically ideal process. The adiabatic curve represents an
adiabatic compression cycle, which is what most conventional
compressor technologies follow. Since the area under the P-V curve
represents the amount of work required for compression, approaching
the isothermal curve means that less work is needed for
compression. A model of one or more compressors according to
various embodiments of the present invention is also shown, nearly
achieving as good of results as the isothermal process. According
to various embodiments, the above-discussed coolant injection
facilitates the near isothermal compression through absorption of
heat by the coolant. Not only does this near-isothermal compression
process require less energy, at the end of the cycle gas
temperatures are much lower than those encountered with traditional
compressors. According to various embodiments, such a reduction in
compressed working fluid temperature eliminates the use of or
reduces the size of expensive and efficiency-robbing
after-coolers.
[0185] Embodiments of the present invention achieve these
near-isothermal results through the above-discussed injection of
liquid coolant. Compression efficiency is improved according to one
or more embodiments because the working fluid is cooled by
injecting liquid directly into the chamber during the compression
cycle. According to various embodiments, the liquid is injected
directly into the area of the compression chamber where the gas is
undergoing compression.
[0186] Rapid heat transfer between the working fluid and the
coolant directly at the point of compression may facilitate high
pressure ratios. That leads to several aspects of various
embodiments of the present invention that may be modified to
improve the heat transfer and raise the pressure ratio.
[0187] One consideration is the heat capacity of the liquid
coolant. The basic heat transfer equation is as follows:
Q=mc.sub.p.DELTA.T
[0188] where Q is the heat, m is mass, .DELTA.T is change in
temperature, and c.sub.p is the specific heat. The higher the
specific heat of the coolant, the more heat transfer that will
occur.
[0189] Choosing a coolant is sometimes more complicated than simply
choosing a liquid with the highest heat capacity possible. Other
factors, such as cost, availability, toxicity, compatibility with
working fluid, and others can also be considered. In addition,
other characteristics of the fluid, such as viscosity, density, and
surface tension affect things like droplet formation which, as will
be discussed below, also affect cooling performance.
[0190] According to various embodiments, water is used as the
cooling liquid for air compression. For methane compression,
various liquid hydrocarbons may be effective coolants, as well as
triethylene glycol.
[0191] Another consideration is the relative velocity of coolant to
the working fluid. Movement of the coolant relative to the working
fluid at the location of compression of the working fluid (which is
the point of heat generation) enhances heat transfer from the
working fluid to the coolant. For example, injecting coolant at the
inlet of a compressor such that the coolant is moving with the
working fluid by the time compression occurs and heat is generated
will cool less effectively than if the coolant is injected in a
direction perpendicular to or counter to the flow of the working
fluid adjacent the location of liquid coolant injection. FIGS.
38A-38D show a schematic of the sequential compression cycle in a
compressor according to an embodiment of the invention. The dotted
arrows in FIG. 38C show the injection locations, directions, and
timing used according to various embodiments of the present
invention to enhance the cooling performance of the system.
[0192] As shown in FIG. 38A, the compression stroke begins with a
maximum working fluid volume (shown in gray) within the compression
chamber. In the illustrated embodiment, the beginning of the
compression stroke occurs when the rotor is at the 6 o'clock
position (in an embodiment in which the gate is disposed at 6
o'clock with the inlet on the left of the gate and the outlet on
the right of the gate as shown in FIGS. 38A-38D). In FIG. 38B,
compression has started, the rotor is at the 9 o'clock position,
and cooling liquid is injected into the compression chamber. In
FIG. 38C, about 50% of the compression stroke has occurred, and the
rotor is disposed at the 12 o'clock position. FIG. 38D illustrates
a position (3 o'clock) in which the compression stroke is nearly
completed (e.g., about 95% complete). Compression is ultimately
completed when the rotor returns to the position shown in FIG.
38A.
[0193] As shown in FIGS. 38B and 38C, dotted arrows illustrate the
timing, location, and direction of the coolant injection.
[0194] According to various embodiments, coolant injection occurs
during only part of the compression cycle. For example, in each
compression cycle/stroke, the coolant injection may begin at or
after the first 10, 20, 30, 40, 50, 60 and/or 70% of the
compression stroke/cycle (the stroke/cycle being measured in terms
of volumetric compression). According to various embodiments, the
coolant injection may end at each nozzle shortly before the rotor
sweeps past the nozzle (e.g., resulting in sequential ending of the
injection at each nozzle (clockwise as illustrated in FIG. 38)).
According to various alternative embodiments, coolant injection
occurs continuously throughout the compression cycle, regardless of
the rotor position.
[0195] As shown in FIGS. 38B and 38C, the nozzles inject the liquid
coolant into the chamber perpendicular to the sweeping direction of
the rotor (i.e., toward the rotor's axis of rotation, in the inward
radial direction relative to the rotor's axis of rotation).
However, according to alternative embodiments, the direction of
injection may be oriented so as to aim more upstream (e.g., at an
acute angle relative to the radial direction such that the coolant
is injected in a partially counter-flow direction relative to the
sweeping direction of the rotor). According to various embodiments,
the acute angle may be anywhere between 0 and 90 degrees toward the
upstream direction relative to the radial line extending from the
rotor's axis of rotation to the injector nozzle. Such an acute
angle may further increase the velocity of the coolant relative to
the surrounding working fluid, thereby further enhancing the heat
transfer.
[0196] A further consideration is the location of the coolant
injection, which is defined by the location at which the nozzles
inject coolant into the compression chamber. As shown in FIGS. 38B
and 38C, coolant injection nozzles are disposed at about 1, 2, 3,
and 4 o'clock. However, additional and/or alternative locations may
be chosen without deviating from the scope of the present
invention. According to various embodiments, the location of
injection is positioned within the compression volume (shown in
gray in FIGS. 38A-38D) that exists during the compressor's highest
rate of compression (in terms of .DELTA.volume/time or
.DELTA.volume/degree-of-rotor-rotation, which may or may not
coincide). In the embodiment illustrated in FIGS. 38A-38D, the
highest rate of compression occurs around where the rotor is
rotating from the 12 o'clock position shown in FIG. 38C to the 3
o'clock position shown in FIG. 38D. This location is dependent on
the compression mechanism being employed and in various embodiments
of the invention may vary. An injection location may also be
selected at an earlier location in the compression chamber (e.g. 9
o'clock in FIGS. 38A-38D to minimize the pressure against which the
liquid must be injected, thus reducing the power required for
coolant injection. Additionally and/or alternatively, liquid (e.g.,
coolant) may be injected into the inlet port before the working
fluid reaches the compression chamber.
[0197] As one skilled in the art could appreciate, the number and
location of the nozzles may be selected based on a variety of
factors. The number of nozzles may be as few as 1 or as many as 256
or more. According to various embodiments, the compressor includes
(a) at least 1, 2, 3, 4, 5, 6, 7, 8, 9, 10, 15, 20, 30, 40, 50, 75,
100, 125, 150, 175, 200, 225, and/or 250 nozzles, (b) less than
400, 300, 275, 250, 225, 200, 175, 150, 125, 100, 75, 50, 40, 30,
20, 15, and/or 10 nozzles, (c) between 1 and 400 nozzles, and/or
(d) any range of nozzles bounded by such numbers of any ranges
therebetween. According to various embodiments, liquid coolant
injection may be avoided altogether such that no nozzles are used.
Along with varying the location along the angle of the rotor
casing, a different number of nozzles may be installed at various
locations along the length of the rotor casing. In certain
embodiments, the same number of nozzles will be placed along the
length of the casing at various angles. In other embodiments,
nozzles may be scattered/staggered at different locations along the
casing's length such that a nozzle at one angle may not have
another nozzle at exactly the same location along the length at
other angles. In various embodiments, a manifold may be used in
which one or more nozzle is installed that connects directly to the
rotor casing, simplifying the installation of multiple nozzles and
the connection of liquid lines to those nozzles.
[0198] Coolant droplet size is a further consideration. Because the
rate of heat transfer is linearly proportional to the surface area
of liquid across which heat transfer can occur, the creation of
smaller droplets via the above-discussed atomizing nozzles improves
cooling by increasing the liquid surface area and allowing heat
transfer to occur more quickly. Reducing the diameter of droplets
of coolant in half (for a given mass) increases the surface area by
a factor of two and thus improves the rate of heat transfer by a
factor of 2. In addition, for small droplets the rate of convection
typically far exceeds the rate of conduction, effectively creating
a constant temperature across the droplet and removing any
temperature gradients. This may result in the full mass of liquid
being used to cool the gas, as opposed to larger droplets where
some mass at the center of the droplet may not contribute to the
cooling effect. Based on that evidence, it appears advantageous to
inject as small of droplets as possible. However, droplets that are
too small, when injected into the high density, high turbulence
region as shown in FIGS. 38B and 38C, run the risk of being swept
up by the working fluid and not continuing to move through the
working fluid and maintain high relative velocity. Small droplets
may also evaporate and lead to deposition of solids on the
compressor's interior surfaces. Other extraneous factors also
affect droplet size decisions, such as power losses of the coolant
being forced through the nozzle and amount of liquid that the
compressor can handle internally.
[0199] According to various embodiments, average droplet sizes of
between 50 and 500 microns, between 50 and 300 microns, between 100
and 150 microns, and/or any ranges within those ranges, may be
fairly effective.
[0200] The mass of the coolant liquid is a further consideration.
As evidenced by the heat equation shown above, more mass (which is
proportional to volume) of coolant will result in more heat
transfer. However, the mass of coolant injected may be balanced
against the amount of liquid that the compressor can accommodate,
as well as extraneous power losses required to handle the higher
mass of coolant. According to various embodiments, between 1 and
100 gallons per minute (gpm), between 3 and 40 gpm, between 5 and
25 gpm, between 7 and 10 gpm, and/or any ranges therebetween may
provide an effective mass flow rate (averaged throughout the
compression stroke despite the non-continuous injection according
to various embodiments). According to various embodiments, the
volumetric flow rate of liquid coolant into the compression chamber
may be at least 1, 2, 3, 4, 5, 6, 7, 8, 9, and/or 10 gpm. According
to various embodiments, flow rate of liquid coolant into the
compression chamber may be less than 100, 80, 60, 50, 40, 30, 25,
20, 15, and/or 10 gpm.
[0201] The nozzle array may be designed for a high flow rate of
greater than 1, 2, 3, 4, 5, 6, 7, 8, 9, 10, and/or 15 gallons per
minute and be capable of extremely small droplet sizes of less than
500 and/or 150 microns or less at a low differential pressure of
less than 400, 300, 200, and/or 100 psi. Two exemplary nozzles are
Spraying Systems Co. Part Number: 1/4HHSJ-SS12007 and Bex Spray
Nozzles Part Number: 1/4YS12007. Other non-limiting nozzles that
may be suitable for use in various embodiments include Spraying
Systems Co. Part Number 1/4LN-SS14 and 1/4LN-SS8. The preferred
flow rate and droplet size ranges will vary with application
parameters. Alternative nozzle styles may also be used. For
example, one embodiment may use micro-perforations in the cylinder
through which to inject liquid, counting on the small size of the
holes to create sufficiently small droplets. Other embodiments may
include various off the shelf or custom designed nozzles which,
when combined into an array, meet the injection requirements
necessary for a given application.
[0202] According to various embodiments, one, several, and/or all
of the above-discussed considerations, and/or
additional/alternative external considerations may be balanced to
optimize the compressor's performance. Although particular examples
are provided, different compressor designs and applications may
result in different values being selected.
[0203] According to various embodiments, the coolant injection
timing, location, and/or direction, and/or other factors, and/or
the higher efficiency of the compressor facilitates higher pressure
ratios. As used herein, the pressure ratio is defined by a ratio of
(1) the absolute inlet pressure of the source working fluid coming
into the compression chamber (upstream pressure) to (2) the
absolute outlet pressure of the compressed working fluid being
expelled from the compression chamber (downstream pressure
downstream from the outlet valve). As a result, the pressure ratio
of the compressor is a function of the downstream vessel (pipeline,
tank, etc.) into which the working fluid is being expelled.
Compressors according to various embodiments of the present
invention would have a 1:1 pressure ratio if the working fluid is
being taken from and expelled into the ambient environment (e.g.,
14.7 psia/14.7 psia). Similarly, the pressure ratio would be about
26:1 (385 psia/14.7 psia) according to various embodiments of the
invention if the working fluid is taken from ambient (14.7 psia
upstream pressure) and expelled into a vessel at 385 psia
(downstream pressure).
[0204] According to various embodiments, the compressor has a
pressure ratio of (1) at least 3:1, 4:1, 5:1, 6:1, 8:1, 10:1, 15:1,
20:1, 25:1, 30:1, 35:1, and/or 40:1 or higher, (2) less than or
equal to 200:1, 150:1, 125:1, 100:1, 90:1, 80:1, 70:1, 60:1, 50:1,
45:1, 40:1, 35:1, and/or 30:1, and (3) any and all combinations of
such upper and lower ratios (e.g., between 10:1 and 200:1, between
15:1 and 100:1, between 15:1 and 80:1, between 15:1 and 50:1,
etc.).
[0205] According to various embodiments, lower pressure ratios
(e.g., between 3:1 and 15:1) may be used for working fluids with
higher liquid content (e.g., with a liquid volume fraction at the
compressor's inlet port of at least 0.5, 1, 2, 3, 4, 5, 6, 7, 8, 9,
10, 15, 20, 25, 30, 35, 40, 50, 60, 70, 75, 80, 85, 90, 91, 92, 93,
94, 95, 96, 97, 98, and/or 99%). Conversely, according to various
embodiments, higher pressure ratios (e.g., above 15:1) may be used
for working fluids with lower liquid content relative to gas
content. However, wetter gases may nonetheless be compressed at
higher pressure ratios and drier gases may be compressed at lower
pressure ratios without deviating from the scope of various
embodiments of the present invention.
[0206] Various embodiments of the invention are suitable for
alternative operation using a variety of different operational
parameters. For example, a single compressor according to one or
more embodiments may be suitable to efficiently compress working
fluids having drastically different liquid volume fractions and at
different pressure ratios. For example, a compressor according to
one or more embodiments is suitable for alternatively (1)
compressing a working fluid with a liquid volume fraction of
between 10 and 50 percent at a pressure ratio of between 3:1 and
15:1, and (2) compressing a working fluid with a liquid volume
fraction of less than 10 percent ata pressure ratio of at least
15:1, 20:1, 30:1, and/or 40:1.
[0207] According to various embodiments, the compressor efficiently
and cost-effectively compresses both wet and dry gas using a high
pressure ratio.
[0208] According to various embodiments, the compressor is capable
of and runs at commercially viable speeds (e.g., between 450 and
1800 rpm). According to various embodiments, the compressor runs at
a speed of (a) at least 350, 400, 450, 500, 550, 600, and/or 650
rpm, (b) less than or equal to 3000, 2500, 2000, 1800, 1700, 1600,
1500, 1400, 1300, 1200, 1100, 1050, 1000, 950, 900, 850, and/or 800
rpm, and/or (c) between 350 and 300 rpm, 450-1800 rpm, and/or any
ranges within these non-limiting upper and lower limits. According
to various embodiments, the compressor is continuously operated at
one or more of these speeds for at least 0.5, 1, 5, 10, 15, 20, 30,
60, 90, 100, 150, 200, 250 300, 350, 400, 450, and/or 500 minutes
and/or at least 10, 20, 24, 48, 72, 100, 200, 300, 400, and/or 500
hours.
[0209] According to various embodiments, the outlet pressure of the
compressed fluid is (1) at least 200, 225, 250, 275, 300, 325, 350,
375, 400, 425, 450, 475, 500, 600, 700, 800, 900, 1000, 1250, 1500,
2000, 3000, 4000, and/or 5000 psig, (2) less than 6000, 5500, 5000,
4000, 3000, 2500, 2250, 2000, 1750, 1500, 1250, 1100, 1000, 900,
800, 700, 600 and/or 500 psig, (3) between 200 and 6000 psig,
between 200 and 5000 psig, and/or (4) within any range between the
upper and lower pressures described above.
[0210] According to various embodiments, the inlet pressure is
ambient pressure in the environment surrounding the compressor
(e.g., 1 atm, 14.7 psia). Alternatively, the inlet pressure could
be close to a vacuum (near 0 psia), or anywhere therebetween.
According to alternative embodiments, the inlet pressure may be (1)
at least -14.5, -10, -5, 0, 5, 10, 25, 50, 100, 150, 200, 250, 300,
350, 400, 450, 500, 550, 600, 700, 800, 900, 1000, 1100, 1200,
1300, 1400, and/or 1500 psig, (2) less than or equal to 3000, 2000,
1900, 1800, 1700, 1600, 1500, 1400, 1300, 1200, 1100, 1000, 900,
800, 700, 600, 500, 400, and/or 350, and/or (3) between -14.5 and
3000 psig, between 0 and 1500 psig, and/or within any range bounded
by any combination of the upper and lower numbers and/or any nested
range within such ranges.
[0211] According to various embodiments, the outlet temperature of
the working fluid when the working fluid is expelled from the
compression chamber exceeds the inlet temperature of the working
fluid when the working fluid enters the compression chamber by (a)
less than 700, 650, 600, 550, 500, 450, 400, 375 350, 325, 300,
275, 250,225, 200, 175, 150, 140, 130, 120, 110, 100, 90, 80, 70,
60, 50, 40, 30, and/or 20 degrees C., (b) at least -10, 0, 10,
and/or 20 degrees C., and/or (c) any combination of ranges between
any two of these upper and lower numbers, including any range
within such ranges.
[0212] According to various embodiments, the outlet temperature of
the working fluid is (a) less than 700, 650, 600, 550, 500, 450,
400, 375, 350, 325, 300, 275, 250, 225, 200, 175, 150, 140, 130,
120, 110, 100, 90, 80, 70, 60, 50, 40, 30, and/or 20 degrees C.,
(b) at least -10, 0, 10, 20, 30, 40, and/or 50 degrees C., and/or
(c) any combination of ranges between any two of these upper and
lower numbers, including any range within such ranges.
[0213] The outlet temperature and/or temperature increase may be a
function of the working fluid. For example, the outlet temperature
and temperature increase may be lower for some working fluids
(e.g., methane) than for other working fluids (e.g., air).
[0214] According to various embodiments, the temperature increase
is correlated to the pressure ratio. According to various
embodiments, the temperature increase is less than 200 degrees C.
for a pressure ratio of 20:1 or less (or between 15:1 and 20:1),
and the temperature increase is less than 300 degrees C. for a
pressure ratio of between 20:1 and 30:1.
[0215] According to various embodiments, the pressure ratio is
between 3:1 and 15:1 for a working fluid with an inlet liquid
volume fraction of over 5%, and the pressure ratio is between 15:1
and 40:1 for a working fluid with an inlet liquid volume fraction
of between 1 and 20%. According to various embodiments, the
pressure ratio is above 15:1 while the outlet pressure is above 250
psig, while the temperature increase is less than 200 degrees C.
According to various embodiments, the pressure ratio is above 25:1
while the outlet pressure is above 250 psig and the temperature
increase is less than 300 degrees C. According to various
embodiments, the pressure ratio is above 15:1 while the outlet
pressure is above 250 psig and the compressor speed is over 450
rpm.
[0216] According to various embodiments, any combination of the
different ranges of different parameters discussed herein (e.g.,
pressure ratio, inlet temperature, outlet temperature, temperature
change, inlet pressure, outlet pressure, pressure change,
compressor speed, coolant injection rate, etc.) may be combined
according to various embodiments of the invention. According to one
or more embodiments, the pressure ratio is anywhere between 3:1 and
200:1 while the operating compressor speed is anywhere between 350
and 3000 rpm while the outlet pressure is between 200 and 6000 psig
while the inlet pressure is between 0 and 3000 psig while the
outlet temperature is between -10 and 650 degrees C. while the
outlet temperature exceeds the inlet temperature by between 0 and
650 degrees C. while the liquid volume fraction of the working
fluid at the compressor inlet is between 1% and 50%.
[0217] According to one or more embodiments, air is compressed from
ambient pressure (14.7 psia) to 385 psia, a pressure ratio of 26:1,
at speeds of 700 rpm with outlet temperatures remaining below 100
degrees C. Similar compression in an adiabatic environment would
reach temperatures of nearly 480 degrees C.
[0218] The operating speed of the illustrated compressor is stated
in terms of rpm because the illustrated compressor is a rotary
compressor. However, other types of compressors may be used in
alternative embodiments of the invention. As those familiar in the
art appreciate, the RPM term also applies to other types of
compressors, including piston compressors whose strokes are linked
to RPM via their crankshaft.
[0219] Numerous cooling liquids may be used. For example, water,
triethylene glycol, and various types of oils and other
hydrocarbons may be used. Ethylene glycol, propylene glycol,
methanol or other alcohols in case phase change characteristics are
desired may be used. Refrigerants such as ammonia and others may
also be used. Further, various additives may be combined with the
cooling liquid to achieve desired characteristics. Along with the
heat transfer and heat absorption properties of the liquid helping
to cool the compression process, vaporization of the liquid may
also be utilized in some embodiments of the design to take
advantage of the large cooling effect due to phase change.
[0220] The effect of liquid coalescence is also addressed in the
preferred embodiments. Liquid accumulation can provide resistance
against the compressing mechanism, eventually resulting in
hydrolock in which all motion of the compressor is stopped, causing
potentially irreparable harm. As is shown in the embodiments of
FIGS. 8 and 17, the inlet 420 and outlet 430 are located at the
bottom of the rotor casing 400 on opposite sides of the gate 600,
thus providing an efficient location for both intake of fluid to be
compressed and exhausting of compressed fluid and the injected
liquid. A valve is not necessary at the inlet 420. The inclusion of
a dwell seal allows the inlet 420 to be an open port, simplifying
the system and reducing inefficiencies associated with inlet
valves. However, if desirable, an inlet valve could also be
incorporated. Additional features may be added at the inlet to
induce turbulence to provide enhanced thermal transfer and other
benefits. Hardened materials may be used at the inlet and other
locations of the compressor to protect against cavitation when
liquid/gas mixtures enter into choke and other cavitation-inducing
conditions.
[0221] Alternative embodiments may include an inlet located at
positions other than shown in the figures. Additionally, multiple
inlets may be located along the periphery of the cylinder. These
could be utilized in isolation or combination to accommodate inlet
streams of varying pressures and flow rates. The inlet ports can
also be enlarged or moved, either automatically or manually, to
vary the displacement of the compressor.
[0222] In these embodiments, multi-phase compression is utilized,
thus the outlet system allows for the passage of both gas and
liquid. Placement of outlet 430 near the bottom of the rotor casing
400 provides for a drain for the liquid. This minimizes the risk of
hydrolock found in other liquid injection compressors. A small
clearance volume allows any liquids that remain within the chamber
to be accommodated. Gravity assists in collecting and eliminating
the excess liquid, preventing liquid accumulation over subsequent
cycles. Additionally, the sweeping motion of the rotor helps to
ensure that most liquid is removed from the compressor during each
compression cycle by guiding the liquid toward the outlet(s) and
out of the compression chamber.
[0223] Compressed gas and liquid can be separated downstream from
the compressor. As discussed below, liquid coolant can then be
cooled and recirculated through the compressor.
[0224] Various of these features enable compressors according to
various embodiments to effectively compress multi-phase fluids
(e.g., a fluid that includes gas and liquid components (sometimes
referred to as "wet gas")) without pre-compression separation of
the gas and liquid phase components of the working fluid. As used
herein, multi-phase fluids have liquid volume fractions at the
compressor inlet port of (a) at least 0.5, 1, 2, 3, 4, 5, 6, 7, 8,
9, 10, 15, 20, 25, 30, 35, 40, 50, 60, 70, 75, 80, 85, 90, 91, 92,
93, 94, 95, 96, 97, 98, 99, and/or 99.5%, (b) less than or equal to
99.5, 99, 98, 97, 96, 95, 94, 93, 92, 91, 90, 85, 80, 75, 70, 60
,50, 40, 35, 30, 25, 20, 15, 10, 9, 8, 7, 6, 5, 4, 3, 2, 1, and/or
0.5%, (c) between 0.5 and 99.5%, and/or (d) within any range
bounded by these upper and lower values.
[0225] Outlet valves allow gas and liquid (i.e., from the wet gas
and/or liquid coolant) to flow out of the compressor once the
desired pressure within the compression chamber is reached. The
outlet valves may increase or maximize the effective orifice area.
Due to the presence of liquid in the working fluid, valves that
minimize or eliminate changes in direction for the outflowing
working fluid are desirable, but not required. This prevents the
hammering effect of liquids as they change direction. Additionally,
it is desirable to minimize clearance volume. Unused valve openings
may be plugged in some applications to further minimize clearance
volume. According to various embodiments, these features improve
the wet gas capabilities of the compressor as well as the
compressor's ability to utilize in-chamber liquid coolant.
[0226] Reed valves may be desirable as outlet valves. As one of
ordinary skill in the art would appreciate, other types of valves
known or as yet unknown may be utilized. Hoerbiger type R, CO, and
Reed valves may be acceptable. Additionally, CT, HDS, CE, CM or
Poppet valves may be considered. Other embodiments may use valves
in other locations in the casing that allow gas to exit once the
gas has reached a given pressure. In such embodiments, various
styles of valves may be used. Passive or directly-actuated valves
may be used and valve controllers may also be implemented.
[0227] In the presently preferred embodiments, the outlet valves
are located near the bottom of the casing and serve to allow
exhausting of liquid and compressed gas from the high pressure
portion. In other embodiments, it may be useful to provide
additional outlet valves located along periphery of main casing in
locations other than near the bottom. Some embodiments may also
benefit from outlets placed on the endplates. In still other
embodiments, it may be desirable to separate the outlet valves into
two types of valves - one predominately for high pressured gas, the
other for liquid drainage. In these embodiments, the two or more
types of valves may be located near each other, or in different
locations.
[0228] The coolant liquid can be removed from the gas stream,
cooled, and recirculated back into the compressor in a closed loop
system. By placing the injector nozzles at locations in the
compression chamber that do not see the full pressure of the
system, the recirculation system may omit an additional pump (and
subsequent efficiency loss) to deliver the atomized droplets.
However, according to alternative embodiments, a pump is utilized
to recirculate the liquid back into the compression chamber via the
injector nozzles. Moreover, the injector nozzles may be disposed at
locations in the compression chamber that see the full pressure of
the system without deviating from the scope of various embodiments
of the present invention.
[0229] According to various embodiments, some compressed working
fluid/gas (e.g., natural gas) that has been compressed by the
compressor is recirculated back into the compression chamber via
the injector nozzles along with coolant to better atomize the
coolant (e.g., similar or identical to how snow-making equipment
combines a liquid water stream with a compressed gas stream to
achieve increase atomization of the water).
[0230] One or more embodiments simplify heat recovery because most
or all of the heat load is in the cooling liquid. According to
various embodiments, heat is not removed from the compressed gas
downstream of the compressor. The cooling liquid may cooled via an
active cooling process (e.g., refrigeration and heat exchangers)
downstream from the compressor. However, according to various
embodiments, heat may additionally be recovered from the compressed
gas (e.g., via heat exchangers) without deviating from the scope of
various embodiments of the present invention.
[0231] As shown in FIGS. 8 and 17, the sealing portion 510 of the
rotor effectively precludes fluid communication between the outlet
and inlet ports by way of the creation of a dwell seal. The
interface between the rotor 500 and gate 600 further precludes
fluid communication between the outlet and inlet ports through use
of a non-contacting seal or tip seal 620. In this way, the
compressor is able to prevent any return and venting of fluid even
when running at low speeds. Existing rotary compressors, when
running at low speeds, have a leakage path from the outlet to the
inlet and thus depend on the speed of rotation to minimize
venting/leakage losses through this flow path.
[0232] The high pressure working fluid exerts a large horizontal
force on the gate 600. Despite the rigidity of the gate struts 210,
this force will cause the gate 600 to bend and press against the
inlet side of the gate casing 152. Specialized coatings that are
very hard and have low coefficients of friction can coat both
surfaces to minimize friction and wear from the sliding of the gate
600 against the gate casing 152. A fluid bearing can also be
utilized. Alternatively, pegs (not shown) can extend from the side
of the gate 600 into gate casing 150 to help support the gate 600
against this horizontal force. Material may also be removed from
the non-pressure side of gate 600 in a non-symmetrical manner to
allow more space for the gate 600 to bend before interfering with
the gate casing 150.
[0233] The large horizontal forces encountered by the gate may also
require additional considerations to reduce sliding friction of the
gate's reciprocating motion. Various types of lubricants, such as
greases or oils may be used. These lubricants may further be
pressurized to help resist the force pressing the gate against the
gate casing. Components may also provide a passive source of
lubrication for sliding parts via lubricant-impregnated or
self-lubricating materials. In the absence of, or in conjunction
with, lubrication, replaceable wear elements may be used on sliding
parts to ensure reliable operation contingent on adherence to
maintenance schedules. These wear elements may also be used to
precisely position the gate within the gate casing. As one of
ordinary skill in the art would appreciate, replaceable wear
elements may also be utilized on various other wear surfaces within
the compressor.
[0234] The compressor structure may be comprised of materials such
as aluminum, carbon steel, stainless steel, titanium, tungsten, or
brass. Materials may be chosen based on corrosion resistance,
strength, density, and cost. Seals may be comprised of polymers,
such as PTFE, HDPE, PEEK.TM., acetal copolymer, etc., graphite,
cast iron, carbon steel, stainless steel, or ceramics. Other
materials known or unknown may be utilized. Coatings may also be
used to enhance material properties.
[0235] As one of ordinary skill in the art can appreciate, various
techniques may be utilized to manufacture and assemble embodiments
of the invention that may affect specific features of the design.
For example, the main casing 110 may be manufactured using a
casting process. In this scenario, the nozzle housings 132, gate
casing 150, or other components may be formed in singularity with
the main casing 110. Similarly, the rotor 500 and drive shaft 140
may be built as a single piece, either due to strength requirements
or chosen manufacturing technique.
[0236] Further benefits may be achieved by utilizing elements
exterior to the compressor envelope. A flywheel may be added to the
drive shaft 140 to smooth the torque curve encountered during the
rotation. A flywheel or other exterior shaft attachment may also be
used to help achieve balanced rotation. Applications requiring
multiple compressors may combine multiple compressors on a single
drive shaft with rotors mounted out of phase to also achieve a
smoothened torque curve. A bell housing or other shaft coupling may
be used to attach the drive shaft to a driving force such as engine
or electric motor to minimize effects of misalignment and increase
torque transfer efficiency. Accessory components such as pumps or
generators may be driven by the drive shaft using belts, direct
couplings, gears, or other transmission mechanisms. Timing gears or
belts may further be utilized to synchronize accessory components
where appropriate.
[0237] After exiting the valves the mix of liquid and gases may be
separated through any of the following methods or a combination
thereof: 1. Interception through the use of a mesh, vanes,
intertwined fibers; 2. Inertial impaction against a surface; 3.
Coalescence against other larger injected droplets; 4. Passing
through a liquid curtain; 5. Bubbling through a liquid reservoir;
6. Brownian motion to aid in coalescence; 7. Change in direction;
8. Centrifugal motion for coalescence into walls and other
structures; 9. Inertia change by rapid deceleration; and 10.
Dehydration through the use of adsorbents or absorbents.
[0238] At the outlet of the compressor, a pulsation chamber may
consist of cylindrical bottles or other cavities and elements, may
be combined with any of the aforementioned separation methods to
achieve pulsation dampening and attenuation as well as primary or
final liquid coalescence. Other methods of separating the liquid
and gases may be used as well.
[0239] FIGS. 39-44 illustrate a compressor 1000 according to an
alternative embodiment. The compressor 1000 is generally similar to
the above-discussed compressors. Accordingly, a redundant
description of similar or identical components is omitted. The
compressor 1000 includes a main casing 1010 that defines a
compression chamber 1020, a drive shaft 1030, a rotor 1040, cams
1050, cam followers 1060, a gate support 1070 (e.g., cam follower
supports, cam struts, gate support arm, gate strut, etc.) connected
to the cam followers 1060, a gate support guide 1075 mounted to the
casing 1010 (or integrally formed with the casing 1010) and
connected to the gate support 1070 to permit reciprocal linear
movement of the gate support 1070, springs 1080 that bias the gate
support 1070 toward the cams 1050, a gate housing 1100 that is
partially formed by and/or mounted to the main casing 1010 and/or
the gate support guide 1075, a gate 1110 slidingly supported by the
gate housing 1100, an inlet manifold 1140 fluidly connected to an
inlet 1150 into the compression chamber 1020, a discharge/outlet
manifold 1160 fluidly connected to a discharge outlet 1170 that
leads from the compression chamber 1020, a discharge outlet valve
1180 disposed in the discharge outlet 1170, coolant injectors 1190,
a hydrostatic bearing arrangement 1300 (see FIGS. 48-51) between
the casing 1010 and gate 1110, and a mechanical/hydraulic seal 1500
that seals the compression chamber 1020 from the ambient
environment around the drive shaft 1030.
[0240] In the illustrated embodiment, the coolant injectors 1190
direct coolant directly into the compression chamber 1020. However,
according to one or more alternative embodiments, coolant
injector(s) 1190 may additionally and/or alternatively inject
coolant into the working fluid in the inlet manifold 1140 before
the working fluid or coolant reach the compression chamber. Such an
alternative may reduce manufacturing costs and/or reduce the amount
of power required to inject the coolant.
[0241] As shown in FIGS. 41, 43, and 44, the discharge outlet valve
1180 directs compressed fluid through the discharge outlet 1170
while discouraging backflow of compressed fluid back into the
compression chamber 1020. As shown in FIG. 41, the valve 1180 is
separately formed from the main casing 1010 and is fitted into the
discharge outlet 1170. However, according to various alternative
embodiments, the valve 1180 or parts thereof may be integrally
formed with the casing 1010.
[0242] As shown in FIGS. 45-46, the discharge manifold 1160
includes a plurality of vanes 1160a. A cross-section of a
passageway within the manifold 1160 from the discharge outlet 1170
(i.e., entrance into the manifold 1160) to a circular discharge
manifold outlet 1160b (i.e., a downstream exit of the manifold
1160) transitions from an axially-elongated cross-section at the
discharge outlet 1170 (e.g., elongated along the length of the gate
1110 in a direction parallel to the rotational axis of the drive
shaft 1030) to the circular discharge manifold outlet 1160b.
According to various embodiments, the cross-sectional area remains
relatively constant throughout this discharge flow path. The vanes
1160a are oriented generally perpendicular to the desired flow path
of the compressed fluid from the compression chamber 1020 to a
discharge manifold outlet 1160b of the discharge manifold 1160. The
vanes 1160a are oriented to promote a generally laminar flow of the
compressed fluid as the cross-sectional shape of the flow path
changes. According to various embodiments, the vanes 1160a reduce
turbulence, increase the efficiency of the compressor 1000, and/or
reduce wear as the compressed fluid (e.g., multiphase liquid/gas
fluid) flows though the outlet 1170 and manifold 1160.
[0243] The vanes 1160a and valve 1180 extend completely across the
flow path of compressed fluid (e.g., into the page as shown in FIG.
45, up and down as shown in FIG. 47, from an upper left toward a
lower right as shown in FIG. 43). The vanes 1160a and valve 1180
therefore structurally support circumferentially-spaced portions
1010a, 1010b (see FIG. 43) of the casing 1010 on either side of the
axially-elongated discharge outlet 1170. The vanes 1160a and valve
1180 may therefore help the casing 1010 to resist deformation
(e.g., that might be encouraged by reaction forces generated
between the gate 1110 and casing 1010 during use of the compressor
1000).
[0244] As shown in FIG. 48, a plurality of vanes/ribs 1155 are
disposed within and extend across the inlet 1150 along the
circumferential direction of the compression chamber 1020 (from
lower left to upper right as shown in FIG. 48). These ribs 1155
strengthen the casing 1010 in the area of the inlet 1150, and help
to prevent deflection of the casing 1010 around the gate 1110.
According to various embodiments, the inlet 1150 is axially divided
into a plurality of discrete inlets 1150 (e.g., holes spaced along
the axial direction of the compressor 1000), such that the
vanes/ribs 1155 are defined by portions of the casing 1010 between
such inlet holes.
[0245] As illustrated in FIGS. 48-51, the compressor 1000 includes
a hydrostatic bearing arrangement 1300 that allows the gate 1110 to
reciprocate up and down relative to the gate housing 1100 while
maintaining close contact with the rotor 1040. The hydrostatic
bearing arrangement 1300 reduces friction between the gate 1110 and
the gate housing 1100.
[0246] As shown in FIGS. 43, 48 and 50, the gate 1110 separates an
inlet side 1020a of the compression chamber 1020 from an outlet
side 1020b of the compression chamber 1020. Pressure in the inlet
side 1020a stays relatively close to the pressure of fluid entering
the compression chamber 1020 via the inlet 1150. Pressure in the
outlet side 1020b of the compression chamber 1020 increases during
each compression stroke/revolution and reaches the output pressure
of compressed fluid being output through the discharge outlet 1170.
As shown in FIG. 50, this causes a higher pressure on the outlet
side 1020b of the gate 1110 than on the inlet side 1020a, which
pushes the gate toward the inlet side 1020a. As shown in FIG. 50,
this differential pressure creates a cantilever force on the gate
1110 and because the compression chamber 1020 pressure increases
until discharge every cycle the cantilever force is constantly
cycling. The hydrostatic bearing arrangement 1300 accommodates this
cycling cantilever force and equalizes the cantilever/bending
moment on the gate 1110.
[0247] As shown in FIGS. 48-51, the hydrostatic bearing arrangement
1300 comprises: upper hydrostatic bearings 1310 on the inlet side
1020a of the gate 1110, lower hydrostatic bearings 1320 on the
inlet side 1020a of the gate 1110, upper hydrostatic bearings 1330
on the compression/outlet side 1020b of the gate 1110, and lower
hydrostatic bearings 1340 on the compression/outlet side 1020b of
the gate 1110.
[0248] As shown in FIG. 49, three of each bearing 1310, 1320, 1330,
1340 are spaced apart along the axial/longitudinal direction of the
compressor 1000 (i.e., into the page as shown in FIG. 50), such
that there are three columns of bearings 1310, 1320, 1330, 1340 (or
six columns if both sides 1020a, 1020b are considered separate).
According to various non-limiting embodiments, the use of multiple
columns of bearings 1310, 1320, 1330, 1340 may reduce the length
the hydraulic fluid has to laterally travel. This may keep
hydraulic fluid more evenly distributed over all surfaces of the
bearing pad. Increasing the number of bearings may also isolate
problems (e.g., debris, deflection of bearing surfaces, wear of
bearing pad surfaces, a clog in the oil system, etc.) to a single
bearing 1310, 1320, 1330, 1340 leaving other bearings 1310, 1320,
1330, 1340 still working properly. However, greater or fewer
columns of bearings 1310, 1320, 1330, 1340 could be used without
deviating from various embodiments (e.g., by combining the
different bearings 1310 into a single longitudinally longer
bearing). According to one or more embodiments, four columns of
bearings are provided on each side of the gate.
[0249] According to various embodiments, the use of multiple
columns of bearings 1310, 1320, 1330, 1340 may facilitate fine
tuning of the resistors 1410 of one column (or bearings within one
column) relative to other column(s) to accommodate for varying
conditions along the length of the gate 1110. For example, if the
hydrostatic pressure causes the sleeve 1360 to bow out in the
middle, the middle column of bearings 1310, 1320, 1330, 1340 can be
tuned down to decrease flow to those larger gaps and increase flow
to the end columns where the gaps are tighter and contact between
the gate and sleeve would first be made.
[0250] As shown in FIGS. 48-50, the hydrostatic bearing arrangement
1300 is formed in a hydrostatic bearing insert/sleeve 1360 that
mates with the casing 1010. Shims or other suitable mechanisms may
be used to ensure a secure, low-tolerance fit and positioning of
the sleeve 1360. The sleeve 1360 is removable from the casing 1010
to facilitate replacement of and/or maintenance on the sleeve 1360.
However, according to alternative embodiments, the insert 1360 may
be integrally formed with the casing 1010.
[0251] As shown in FIG. 51, each bearing 1310, 1320, 1330, 1340
comprises an inlet port 1310a, 1320a, 1330a, 1340a that opens into
a pocket groove 1310b, 1320b, 1330b, 1340b on a side of the insert
1360 that mates with the gate 1110. Each groove 1310b, 1320b,
1330b, 1340b is surrounded by a land/bearing pad 1310c, 1320c,
1330c, 1340c that closely mates with the gate 1110. The pad 1310c,
1320c, 1330c, 1340c is surrounded by a drain 1370, which may be
common to all of the bearings 1310, 1320, 1330, 1340.
[0252] As shown in FIG. 51, a hydraulic pump 1380 pumps hydraulic
fluid (e.g., oil) from a reservoir 1390 through hydraulic
passageways 1400 to respective resistor flow valves 1410 for each
of the bearings 1310, 1320, 1330, 1340. The passageways 1400 then
lead sequentially to respective inlet ports 1310a, 1320a, 1330a,
1340a, grooves 1310b, 1320b, 1330b, 1340b, lands/bearing pads
1310c, 1320c, 1330c, 1340c, the drain 1370, and back into the
reservoir 1390.
[0253] As already known, hydrostatic bearings work by using two
flow resistors. In this embodiment, the first flow resistor is a
flow resistor valve 1410 inline prior to the bearing 1310, 1320,
1330, 1340, which is held constant during operation. The bearing
pad 1310c, 1320c, 1330c, 1340c itself is the second flow resistor.
The resistance of the bearing pad 1310c, 1320c, 1330c, 1340c
changes and is dependent on the gap between the gate 1110 and the
bearing pad itself 1310c, 1320c, 1330c, 1340c. If this gap
decreases the pressure in the bearing pad 1310c, 1320c, 1330c,
1340c and the pocket grooves 1310b, 1320b, 1330b, 1340b will go up
and similarly if the gap increases the pressure in the pad 1310c,
1320c, 1330c, 1340c and the pocket grooves 1310b, 1320b, 1330b,
1340b will go down. The gap will change due to loads created by the
cantilever pressure force on the gate 1110.
[0254] According to various embodiments, the flow resistor valve
1410 can be replaced by a set flow resistor or an annulus in the
respective passageway 1400 that behaves similarly to the bearing
pad resistor. An annulus can be designed into the bearing pad
1310c, 1320c, 1330c, 1340c that allows flow to pass through it with
a resistance that is dependent on the gap. Typically the annulus is
placed on the opposite surface of the bearing pad to which it is
hydraulically connected. To be clear, lubricant would flow through
the annulus on one side of the bearing and then flow to its
respective bearing pad on the opposite side. Thus, according to
various embodiments, the bearings 1310, 1320, 1330, 1340 comprise
self-compensating bearings with flow resistors built into the
opposing bearings. For example, the flow resistor valve 1400 for
the bearing 1310 may be built into the opposite bearing 1330 so
that flow to the bearing 1310 is reduced when the bearing 1330 gap
is reduced. This may prevent excess hydraulic fluid flow through
bearings 1310, 1320, 1330, 1340 with large gaps (because the gap on
the opposing bearing is small) or permit larger flow rates to
bearings 1310, 1320, 1330, 1340 that have higher loads. Bearings
1320, 1340 oppose each other and can work in the same manner. This
type of self-compensating hydrostatic bearing is described in U.S.
Pat. No. 7,287,906, the entire contents of which are incorporated
herein by reference.
[0255] As shown in FIG. 50, according to various embodiments, the
use of upper bearings 1310, 1330 that are discrete from lower
bearings 1320, 1340 enables the bearing arrangement 1300 to adapt
to the cantilever/bending moments being exerted on the gate 1110 by
the pressurized fluid in the compression chamber 1020, 1020b and
the rotor 1040. The magnitude of the forces being exerted on the
gate 1110 by the inlet and outlet sides 1020a, 1020b of the
compression chamber 1020 and the bearings 1310, 1320, 1330, 1340 is
represented by the size of the arrows. As shown in FIG. 50, when
the outlet side 1020b force is high relative to the inlet side
1020a, the moment is balanced by a high force from the upper
far-side bearing 1310 and lower near-side bearing 1340, where the
gaps are the smallest. Conversely, the bearing gaps are larger
between the gate 1110 and bearings 1320, 1330, such that the force
applied by these bearings 1320, 1330 is lower. According to various
alternative embodiments, additional upper, lower, and/or
intermediate hydrostatic bearings may be added to more specifically
account for the bending moment being exerted on the gate 1110.
However, according to alternative embodiments, the upper and lower
hydrostatic bearings (e.g., bearings 1330, 1340; bearings 1310,
1320) may be combined without deviating from the scope of various
embodiments.
[0256] As used herein, the directional terms "upper" and "lower"
with respect to bearings 1310, 1330, 1320, 1340 are defined along
the direction of reciprocating movement of the gate 1110, and not
necessarily along a gravitational up/down direction (though
gravitational up/down aligns with the gate 1110's up/down
reciprocating direction according to various embodiments).
[0257] According to various embodiments, the hydrostatic bearing
arrangement 1300 creates a fluid film gap between the gate 1110 and
casing 1010 on the inlet side 1020a of the compression chamber
1020, which may prolong the useful life of the gate 1110 and/or
casing 1010 by reducing or eliminating wearing contact between the
gate 1110 and casing 1010, and/or reduce the forces required to
move the gate 1110 along its reciprocating path.
[0258] According to various alternative embodiments, the
hydrostatic bearing is used on a rotary vane compressor in which
the vanes rotate with and reciprocate relative to the rotor instead
of the casing. In such embodiments, a hydrostatic bearing such as
the bearing 1300 is disposed between the rotor and gate, rather
than between the casing and gate.
[0259] As shown in FIG. 50, the gate 1110 includes a seal 1430 that
mounts to a groove 1440a in the main body 1440 of the gate 1110. As
shown in FIG. 50, the seal 1430 and groove 1440a have complimentary
"+" shaped profiles that help to retain the seal 1430 in the groove
1440a during operation of the compressor 1000. According to various
alternatives, the groove 1440a and seal 1430 may have any other
suitable complimentary profile that discourages separation of the
seal 1430 from the gate body 1440 (e.g., a profile with a narrow
top opening and a larger (e.g., bulbous) middle cross-section, a
triangular profile with a point toward the top, etc.).
[0260] As shown in FIG. 50, according to various embodiments, the
gate body 1440 and/or the sleeve 1360 may be formed from hard
materials that resist wear (e.g., materials such as 440C steel,
17-4 steel, D2 tool steel, or Inconel, among others, with HRC over
35, 40, 45, 50, 55, 60, 65,etc.) or are coated with wear-resistant
coatings or otherwise treated to increase hardness (e.g., nitrided
steel, steel with a hard ceramic coating, steel with surface heat
treatments that increase surface hardness, etc.) so as to resist
wear when and if the sleeve 1360 and gate body 1440 rub against
each other. Additionally and/or alternatively, one of the sleeve
1360 and gate body 1440 may have a hard surface (e.g., steel) while
the other of the sleeve 1360 and gate body 1440 is relatively
softer (e.g., formed of bronze of brass) so as to be sacrificially
worn during operation, and eventually replaced. According to one or
more embodiments, the sleeve 1360 comprises a hard-surfaced
material such as steel, while the gate body 1440 comprises a soft
material such as bronze. According to one or more alternative
embodiments, the sleeve 1360 comprises a soft material such as
bronze, while the gate body 1440 comprises a hard material such as
steel.
[0261] According to various embodiments, the surface of the gate
1110 and/or sleeve 1360 (or a coating thereon) is matted or
otherwise constructed so as to create turbulence within the oil
flow, thereby increasing the shear force of the oil as it forces
its way through the gaps and increases the hydrostatic bearing
pressure.
[0262] According to alternative embodiments, the hydrostatic
bearing arrangement 1300 is replaced with a hydrodynamic bearing
arrangement, which provides hydraulic liquid (e.g., oil) to an
interface between the gate body 1440 and sleeve 1360. The
hydrodynamic bearing relies on relative movement between the gate
body 1440 and sleeve 1360 to cause the hydraulic fluid to
pressurize and/or lubricate the intersection.
[0263] As shown in FIG. 40, a mechanical seal 1500 on each axial
end of the compressor 1000 hermitically seals the compression
chamber 1020 of the compressor 1000 relative to the environment
outside of the compression chamber 1020 around the driveshaft
1030.
[0264] Each of the two mechanical seals 1500 includes face seals
1510, 1520, a radial shaft seal 1550, a vent 1560, and hydraulic
packing 1590. As shown in FIGS. 40, 52, and 54, the inner and outer
face seals 1510, 1520 seal an axial end of the rotor 1040 relative
to the axial face of the casing 1010 that defines the compression
chamber 1020. As shown in FIG. 52, the seals 1510, 1520 are mounted
within circumferential (but non-circular in the case of seal 1520)
face grooves 1040b in the rotor 1040 to permit axial movement
(i.e., left/right movement as shown in FIG. 40), and springs 1530,
1540 (e.g., Belleville washers, an O-ring with elastic properties,
a series of compression springs arranged around the perimeter of
the seals 1501, 1520) bias the seals 1510, 1520 axially against the
axial face of the casing 1010 that defines the compression chamber
1020. The inner face seal 1510 is circular and concentric with a
rotational axis of the drive shaft 1030. As shown in FIG. 41, the
outer face seal 1520 follows the non-circular perimeter of the
rotor 1040, and rotates with the rotor 1040 about the axis of the
drive shaft 1030. According to various embodiments, outer sealing
portions of the face seals 1510, 1520 comprise low-friction
material (e.g., graphite) that is bonded to a stronger backing
(e.g., steel).
[0265] According to various embodiments, the seals 1510, 1520 are
retained in their grooves 1040b even when the wear surface of the
seals 1510, 1520 (e.g., the graphite portion of the seals 1510,
1520) is worn through. For example, as shown in FIGS. 67 and 68,
the seals 1510, 1520 may be retained by locking washers 1541 (e.g.,
multiple washers per seal 1510, 1520) that are connected (e.g., via
bolts 1542 or other fasteners) to recesses 1040c in the end faces
of the rotor 1040 and extend into shouldered grooves 1510a, 1520a
in the seals 1510, 1520 to prevent the seals 1510, 1520 from
separating from mating seal grooves 1040b, while permitting the
seals 1510, 1520 to move axially within the grooves 1040b to keep
the seals 1510, 1520 proximate to the mating face of the
compression chamber (e.g., the face of wear plate 1545 (see FIG.
52).
[0266] As shown in FIG. 52, an end cap wear plate 1545 on each
axial end of the compression chamber 1020 removably mounts to a
remainder of the casing 1010 (e.g., via bolts) and abuts the seals
1510, 1520. The plate 1545 may be replaced when wearing contact
between the seals 1510, 1520 and plate 1545 has worn the plate 1545
sufficiently to warrant replacement.
[0267] As shown in FIG. 54, the radial shaft seal 1550 extends
radially between the drive shaft 1030 and an end cap of the casing
1010. As shown in FIGS. 54 and 40, the vent 1560 is disposed
axially outwardly from the radial shaft seals 1550. As shown in
FIG. 54, a fluid passageway 1570 fluidly connects the vent 1560 to
the inlet 1150 of the compressor 1000. As shown in FIG. 54, the
hydraulic packing 1590 comprises facing radial seals 1600, 1610
with a hydraulic fluid passage 1620 therebetween. The hydraulic
pump 1380 (or any other suitable source of hydraulic fluid)
provides pressurized hydraulic fluid to the hydraulic packing 1590
via a port/passageway 1630 that leads into the space between the
seals 1600, 1610. As shown in FIG. 54, rotational bearings 1650
support the drive shaft 1030 relative to the casing 1010 to permit
the drive shaft 1030 to rotate relative to the casing 1010.
[0268] The operation of the mechanical seal 1500 is described with
reference to FIGS. 52 and 54. For the working fluid (e.g., natural
gas being compressed) to leak out of the compression chamber 1020,
the fluid may leak sequentially through the seals 1520, 1510, 1550.
If the working fluid leaks past all three seals 1520, 1510, 1550,
the fluid reaches the vent 1560 which returns the fluid back to the
compressor inlet 1150 via the passageway/port 1570, which is
maintained at the pressure of the inlet 1150 via its fluid
communication with the inlet 1150. The hydraulic packing 1590 on
the outer axial side of the vent 1560 is pressurized via hydraulic
fluid to a pressure higher than the inlet 1150 pressure, which
discourages or prevents the working fluid from further leaking past
the hydraulic packing 1590. Leaked working fluid leaks through the
passageway/port 1570 back to the intake 1150, rather than past the
hydraulic packing 1590 because the inlet 1150 is at a significantly
lower pressure than the hydraulic packing 1590. Thus, leakage of
the working fluid past the hydraulic packing 1590 is reduced or
preferably eliminated. Pressure in the bearing cavity for the
bearings 1650 is maintained at ambient atmospheric pressure.
[0269] According to various embodiments, the mechanical seal 1500
provides an axially-compact seal that results in lower moment loads
on the compressor's bearings.
[0270] As shown in FIG. 52, in the compressor 1000, the drive shaft
1030 is mounted to each axial end of the casing 1010 via a
combination of separate rotational bearings 1650 and thrust
bearings 1660. However, as shown in FIG. 53, the separate
rotational and thrust bearings 1650, 1660 may be replaced by a
consolidated bearing 1670 that serves both thrust bearing and
rotational bearing functions without deviating from the scope of
various embodiments. To facilitate removal of the bearing 1670 from
the drive shaft, a lubrication passageway may extend through the
drive shaft and open into the interface between the drive shaft and
the bearing 1670. According to various alternative embodiments, the
bearings 1650, 1660 may be replaced with any other type of
rotational coupling between the drive shaft 1030 and casing 1010
without deviating from the scope of various embodiments (e.g.,
other types of bearings, bushings, etc.).
[0271] Although the seal 1500 is described as including various
structures in the illustrated embodiment, the seal 1500 may include
greater or fewer structures without deviating from the scope of the
present invention. For example, one or more of the seals 1510,
1520, 1550 may be omitted without deviating from the scope of the
present invention.
[0272] FIG. 69 illustrates a compressor 5150 that is generally
similar to the compressor 1000, except that the compressor 5150
uses an alternative embodiment of a mechanical seal 5200 in place
of the mechanical seal 1500. The mechanical seal 5200 is generally
similar to the seal 1500, so a redundant explanation of similar or
identical components is omitted. In contrast with the axially
spaced arrangement of various components of the mechanical seal
1500 (e.g., the radial seal 1550, vent 1560, radial seals 1600,
1610, and pressurized hydraulic fluid passageway 1620), various
components of the mechanical seal 5200 are radially spaced from
each other, which may provide a more axially-compact seal. As shown
in FIG. 69, the compressor 5150 includes a casing 5210 that is
generally identical to the casing 1010, except that the casing 5210
is shaped slightly differently so as to accommodate the differently
shaped mechanical seal 5200.
[0273] As shown in FIG. 69, the seal 5200 includes an annular
collar 5220 that is rigidly and sealingly connected to or
integrally formed with the drive shaft 1030 so as to rotate with
the drive shaft 1030 relative to the casing 5210. According to
various embodiments, the collar 5220 may connect to the drive shaft
1030 in a variety of alternative ways (e.g., heat-shrunk onto the
shaft 1030, glued or otherwise fastened onto the shaft 1030, welded
onto the shaft 1030, press-fit onto the shaft 1030, etc.).
According to various embodiments, o-rings 5230 are disposed between
the collar 5220 and shaft 1030 to prevent leaks therebetween. Inner
annular seal grooves 5220a,b and outer annular seal grooves 5220c,d
are disposed on the axial faces of the collar 5220 that face toward
and away from the rotor 1040. Face seals 5240, 5250, 5260, 5270 are
disposed in the grooves 5220a,b,c,d and spring biased away from the
collar 5220 toward a mating axial face surface 5210a,5210b of the
casing 5210. A vent 5290 is disposed between the collar 5220 and
casing 5210 radially outwardly from the collar 5220. The vent 5290
fluidly connects to an inlet into the compressor 5150 via a
passageway 5300 in the casing 5210. A hydraulic fluid passageway
5310 connects a source of pressurized hydraulic fluid (or other
fluid) (e.g., the pump 1380) to a space 5330 disposed between the
seals 5250, 5270, face 5210b, and collar 5220 so as to keep this
space 5330 pressurized with hydraulic fluid.
[0274] The operation of the mechanical seal 5200 is described with
reference to FIG. 69. If working fluid leaks from the compression
chamber 1020 sequentially past the face seal 1520, face seal 1510,
face seal 5240, and face seal 5260, the leaked working fluid will
leak into the vent 5290, which will direct the leaked working fluid
back to the inlet of the compressor 5150 via the passageway 5300.
As with the seal 1500, the hydraulic packing formed by the seals
5250, 5270, and the pressurized fluid disposed in the space 5330
discourages or prevents leaked working fluid in the vent 5290 from
further leaking past the seals 5250, 5270. Because the pressure in
the inlet into the compressor 5150 is lower than the pressure in
the space 5330, leaked fluid will flow back to the inlet rather
than leaking past the hydraulic packing.
[0275] According to various embodiments, the seal 5200 may be
modified by adding or removing various seals. For example, the
compressor 5150 includes one more seal between the compression
chamber and the vent than is included in the compressor 1000. In
particular, in the compressor 5150, four seals are disposed between
the compression chamber 1020 and the vent 5290 (i.e., the seals
1520, 1510, 5240, 5260), while the illustrated compressor 1000 has
three such seals (i.e., seals 1520, 1510, 1550). However, according
to alternative embodiments greater or fewer such seals may be
disposed between the compression chamber and vent without deviating
from the scope of various embodiments. For example, one or more of
the seals 1520, 1510, 5240, 5260 may be omitted. Alternatively,
additional seals like the seals 5240, 5260 may extend between the
collar 5220 and the face 5210a of the casing 5210 to further reduce
leakage from the compression chamber 1020, and the collar 5220 and
faces 5210a,b may be radially expanded to provide space for such
additional seals, preferably without axially elongating the overall
mechanical seal. Additionally and/or alternatively, the seal 5200
may be modified by adding a radial seal (e.g., like the seal 1550)
between the casing 5210 and shaft 1030 along the leakage path
between the seals 1510, 5240. Additionally and/or alternatively,
the vent 5290 may be disposed along the leakage path between
different ones of the seals 1520, 1510, 5240, 5260. For example,
the vent may alternatively be disposed in the leakage path between
the inner face seal 5240 and the outer face seal 5260.
[0276] As shown in FIGS. 41 and 43, according to various
embodiments, one or more holes 1040a extend axially through the
entire rotor 1040 so as to fluidly connect opposite axial ends of
the rotor 1040 radially inwardly from the seals 1520. These holes
1040a may prevent the rotor 1040 from being axially pushed against
one axial end of the compression chamber 1020 if compressed working
fluid asymmetrically leaked past one of the seals 1520 on one axial
end of the rotor 1040 to a greater extent than at the opposite
axial end of the rotor 1040. Additionally and/or alternatively, the
fluid communication between the axial ends of the rotor 1040 may be
provided by extending a fluid passageway through the end plates
1545 of the casing 1010 (see FIG. 52), instead of through the rotor
1040.
[0277] As shown in FIG. 52, according to various embodiments, a
proximity sensor 1580 (e.g., contact or non-contact sensor,
capacitive sensor, magnetic sensor, etc.) monitors the axial
position of the rotor 1040 relative to the end plates 1545 or other
part of the casing 1010. The sensor 1580 and associated controller
(e.g., electronic control unit, analog or digital circuitry, a
computer such as a PC, etc.) may cause one or more actions (e.g.,
an audio or visual alarm, deactivation of the compressor) to occur
if the sensed distance exceeds a predetermined distance or falls
below a predetermined distance
[0278] FIGS. 55-58 illustrate a compressor 2000 according to an
alternative embodiment. The compressor 2000 is generally similar to
the above-discussed compressors. Accordingly, a redundant
description of similar or identical components is omitted. The
compressor 2000 includes a main casing 2010 that defines a
compression chamber 2020, a drive shaft 2030, a rotor 2040 mounted
to the drive shaft 2030 for rotation with the drive shaft 2030
relative to the casing 2010, a gate 2050 slidingly connected to the
casing 2010 for reciprocating movement, and a gate-positioning
system 2060. The gate-positioning system 2060 of the compressor
2000 differs from the gate-positioning systems of the
above-described compressors.
[0279] As shown in FIGS. 55-58, the gate-positioning system 2060
includes: a gate-positioning-system casing 2070 mounted to the main
casing 2010 (e.g., via bolts or integral formation) (see FIGS. 56
and 58), a drive pulley 2080 mounted to the driveshaft 2030 for
rotation with the drive shaft 2030, a cam shaft 2090 rotationally
mounted to the casing 2070 for relative rotation about a cam shaft
axis that is parallel to an axis of the main drive shaft 2030, a
driven pulley 2095 mounted to the cam shaft 2090 for rotation with
the cam shaft 2090 relative to the casings 2070, 2010, a belt 2100
connected to the pulleys 2080, 2095, two cams 2110 mounted to the
camshaft 2090 for rotation with the camshaft 2090, cam followers
2120 rotationally mounted to gate supports 2130 for rotation
relative to the supports 2130 about axes that are parallel to the
rotational axes of the shafts 2030, 2090, and springs 2140 that
extend between the casing(s) 2070, 2010 and the gate supports
2130.
[0280] The gate supports 2130 mount to the gate 2050 to drive the
reciprocating motion of the gate 2050. As shown in FIG. 57, the
gate supports 2130 pass through enlarged lower openings 2050a in
the gate 2050 and rigidly attach (e.g., via a threaded connection,
a retainer key or ring, a retainer pin 2135 (as shown in FIG. 57),
etc.) to upper portions of the gate 2050 near an upper sealing edge
2050b of the gate 2050. The lower openings 2050a are enlarged
relative to the gate supports 2130 so that the gate supports 2130
do not contact lower portions of the gate 2050. According to
various embodiments, extending the gate supports 2130 through the
enlarged lower openings 2050a limits the effect that thermal
expansion/contraction has on the positioning of the seal 2050b of
the gate 2050 relative to the gate support 2130 position. In
particular, thermal expansion of the gate 2050 below where the gate
2050 mounts to the gate supports 2130 does not affect the
positioning of the gate's seal 2050b relative to the gate supports
2130. According to various embodiments, this provides more precise
and accurate gate seal 2050b positioning relative to the rotor 2040
when the gate 2050 thermally expands or contracts during use of the
compressor 2000.
[0281] As shown in FIGS. 56 and 57, the gate supports 2130
slidingly mount to the casing 2070 and/or 2010 via linear bearings
2137 (or other linear connections such as bushings, etc.) to permit
the gate supports 2130 to move in the reciprocating direction of
the gate 2050 (up/down as shown in FIGS. 56 and 57). An upper end
of the springs 2140 abuts a spring retainer portion of the casing
2070 and/or casing 2010. A lower end of the springs 2140 connects
to the gate supports 2130 via spring retainers 2150 or other
suitable connectors. As a result, the compression springs 2140 urge
the gate supports 2130 and gate 2050 downwardly away from the rotor
2040 and towards the cams 2110.
[0282] During operation of the compressor 2000, the drive shaft
2030 rotationally drives the pulley 2080, which rotationally drives
the belt 2100, which rotationally drives the pulley 2095, which
rotationally drives the shaft 2090, which rotationally drives the
cams 2110. Rotation of the cams 2110 drives the cam followers 2120,
gate support 2130, and gate 2050 upwardly toward the rotor 2040
against the spring bias of the springs 2140. The cams 2110 are
shaped and the belt 2100 and pulleys 2080, 2095 are timed so that
the gate positioning system 2060 maintains the seal 2050b of the
gate 2050 proximate to (e.g., within 5, 4, 3, 2, 1, 0.5, 0.3, 0.1,
0.05, 0.04, 0.03, 0.02, 0.01, 0.005, 0.004, 0.003, 0.002, and/or
0.001 mm of) the rotor 2040 as the rotor 2040 rotates during
operation of the compressor 2000. The gate-positioning system 2060
therefore generally works in a similar manner as the gate
positioning system illustrated in FIG. 1, except that the relative
roles of the springs and cams are reversed in the compressor 2000
(i.e., the cams 2110 urge the gate 2050 toward the rotor 2040,
rather than away from it, and the springs 2140 urge the gate 2050
away from the rotor 2040, rather than toward it).
[0283] In the gate-positioning system 2060 according to various
non-limiting embodiments, a mass of the reciprocating components
(e.g., the gate 2050, gate supports 2130, cam followers 2120,
portions of the springs 2140 and retainers 2150) is kept relatively
low to reduce the forces needed to drive such reciprocation.
According to various embodiments, such reduction in reciprocating
mass may facilitate higher compressor 2000 operational speeds (in
terms of RPMs) and/or smaller springs 2140 and other structural
components of the system 2060.
[0284] In the illustrated embodiment, the cam shaft 2090 is
belt-driven via the pulleys 2080, 2095 and belt 2100. However,
according to alternative embodiments, the cam shaft 2090 may be
driven by any other suitable mechanism for transferring rotation
from the drive shaft 2030 to the cam shaft 2090 (e.g., chain drive,
gear drive, etc.) without deviating from the scope of various
embodiments.
[0285] As shown in FIGS. 56-58, the casing 2070 encloses many of
the components of the gate-positioning system 2060. In the
illustrated embodiment, the only working fluid leakage path to the
ambient environment via the gate 2050/casing 2010 interface is via
the intersection between a hole 2070a in the casing 2070 and the
cam shaft 2090 on the side of the casing 2070 where the cam shaft
2090 projects through the casing 2070 so that it may be driven by
the pulley 2095. As shown in FIG. 57, a hydraulic packing 2170
seals this leakage path/intersection between the cam shaft 2090 and
casing 2070. According to various embodiments, the hydraulic
packing 2170 may be similar to or identical to the above-discussed
hydraulic packing 1590, and may comprise facing radial seals (e.g.,
similar to or identical to the seals 1600, 1610) with a hydraulic
fluid passage (e.g., similar to or identical to the passage 1620)
therebetween. The hydraulic pump 1380 may provide pressurized
hydraulic fluid to the hydraulic packing 2170 via a port/passageway
(e.g., similar to or identical to the port/passageway 1630) that
leads into the space between the seals. As a result, the pressure
within the hydraulic packing 2170 exceeds a pressure within the
casing 2070 so that fluids (e.g., working fluid that leaked past
the gate 2050 into the casing 2070 volume) do not leak out of or
are discouraged from leaking out of the casing 2070. The casing
2070 may be pressurized by working fluid that escaped from the
compression chamber 2020, and that pressure may prevent or
discourage further leakage through that flow path.
[0286] Additionally and/or alternatively, as shown in FIG. 56, a
vent passage 2180 may fluidly connect the interior of the casing
2070 with the inlet (e.g., via the inlet manifold 2190 or a direct
connection to the inlet in the casing 2010). Such a vent passage
2180 may help to ensure that a pressure in the casing 2070 remains
below a hydraulic pressure in the hydraulic packing 2170 so as to
further discourage working fluid in the casing 2070 from leaking
past the hydraulic packing 2170.
[0287] According to alternative embodiments, the hydraulic packing
2170 may be replaced with any other suitable seal (e.g.,
conventional hermetic seals that are designed to seal rotating
shafts where there is a significant pressure differential between
opposing sides of the seal) or eliminated altogether (e.g., if the
gate 2050's seal is sufficient) without deviating from the scope of
various embodiments.
[0288] According to an alternative embodiment, the casing 1010 and
2070 are axially extended to entirely enclose the pulleys 2080,
2095 and cam shaft 2090 such that only the main drive shaft 2030 of
the compressor 2000 extends from the casing 2010, 2070, requiring a
single mechanical seal like the seal 2170 between the drive shaft
2030 and elongated casing to hermetically seal the compressor
2000.
[0289] FIGS. 59-60 illustrate a compressor 3000 according to an
alternative embodiment. The compressor 3000 is generally similar to
the above-discussed compressor 2000. Accordingly, a redundant
description of similar or identical components is omitted. The
compressor 3000 differs from the compressor 2000 by adding two
additional sub-compressors that are axially spaced from each other.
Thus, the compressor 3000 comprises three sub-compressors 3000a,
3000b, 3000c. The compressor 3000 includes a main casing 3010 that
defines three compression chambers 3020a, 3020b, 3020c, a drive
shaft 3030, three rotors 3040a, 3040b, 3040c mounted to the drive
shaft 3030 for rotation with the drive shaft 3030 relative to the
casing 3010, three gates 3050a, 3050b, 3050c slidingly connected to
the casing 3010 for reciprocating movement, and a gate-positioning
system 3060 that includes three cams 3110a, 3110b, 3110c mounted to
the cam shaft 3090, three cam followers 3120a, 3120b, 3120c, three
gate supports 3130a, 3130b, 3130c, and three springs 3140a, 3140b,
3140c. The gate-positioning system 2060 of the compressor 2000
differs from the gate-positioning systems of the above-described
compressors. Each of the respective sets of a, b, and c components
(e.g., compression chamber 3020a, rotor 3040a, gate 3050a, cam
3110a, cam follower 3120a, gate support 3130a, and spring 3140a)
work in substantially the same manner as the comparable components
of the whole compressor 2000.
[0290] The inlet manifold 3500 of the compressor 3000 fluidly
connects to the inlets of each sub-compressor 3000a, 3000b, 3000c.
According to various embodiments, the working fluid inlets of the
three sub-compressors 3000a, 3000b, 3000c fluidly connect to each
downstream from the manifold 3500. Similarly, the compressed
working fluid outlets of the three sub-compressors 3000a, 3000b,
3000c rejoin in the compressor's discharge manifold 3510. According
to various embodiments, check-valves are disposed in each
sub-compressor's discharge outlets upstream from where the
discharge passageways join together.
[0291] According to various embodiments, check-valves are also
disposed in each sub-compressor's inlet downstream from where the
inlet flow path diverges toward respective sub-compressors 3000a,
3000b, 3000c (e.g., downstream or within the inlet manifold 3500)
so as to discourage backflow from one chamber 3020a, 3020b, 3020c
into another chamber 3020a, 3020b, 3020c during out-of-phase
operation of the sub-compressors 3000a, 3000b, 3000c.
[0292] As shown in FIGS. 59 and 60, the compression cycles of the
compressors 3000a, 3000b, 3000c are 120 degrees out of phase with
each other. Thus, when the sub-compressor 3000a begins its
compression cycle, the sub-compressor 3000b is 1/3 of the way
through its cycle, and the sub-compressor 3000c is 2/3 of the way
through its cycle. Positioning the sub-compressors 3000a, 3000b,
3000c out of phase in this manner reduces the maximum instantaneous
torque that must be applied to the compressor 3000, which may
reduce the size/power/HP of the engine, motor, or other rotational
driver being used to drive the drive shaft 3030 of the compressor
3000. The 3-phase operation of the compressor 3000 may also reduce
vibrations as the reciprocating movement of the gate-positioning
system are generally balanced across the three sub-compressors
3000a, 3000b, 3000c. The 3-phase operation of the compressor 3000
may also reduce pressure spikes downstream from the compressor 3000
(e.g., in the discharge manifold 3510) because the compressed fluid
flow is divided into three sequential bursts for each revolution of
the drive shaft 3030 (as opposed to a single larger burst in the
compressor 2000). The 3-phase operation of the compressor 3000 may
also increase the strength of the casing 3010 and reduce the
required reinforcement of the casing 3010 around the gate because
the single gate slot of the compressor 2000 is replaced with 3 gate
slots with reinforcing structure therebetween. The 3-phase
operation of the compressor 3000 may reduce the cost of the
compressor 3000 because the narrower gates 3050a, 3050b, 3050c or
rotors 3040a, 3040b, 3040c (or other components of the compressor
3000) may be more easily fabricated because they are not as long.
The 3-phase operation of the compressor 3000 may reduce the cost of
the compressor 3000 because bearings may be disposed between
adjacent compression chambers 3020a, 3020b, 3020c, which can reduce
drive shaft 3030 deflection, and facilitate less expensive drive
shafts 3030 and other components, while still maintaining tight
tolerances between the rotor 3040a, 3040b, 3040c and casing
3010.
[0293] While the illustrated compressor 3000 includes three
sub-compressors 3000a, 3000b, 3000c, the compressor may include
greater or fewer sub-compressors without deviating from the scope
of various embodiments (e.g., n sub-compressors that operate out of
phase by 360/n degrees from each other, where n is an integer
greater than 1 and preferably less than 100 (e.g., 2, 3, 4, 5, 6,
7, 8, 9, 10)).
[0294] Alternatively, the multi-phase concept of the compressor
3000 may be implemented using three discrete compressors (e.g., any
of the above discussed compressors such as the compressors 1000,
2000, 5150) by connecting their respective drive shafts (e.g., via
direct co-axial mounting such that the compressors are axially
spaced from each other along a common drive shaft, via gears,
belts, etc.) such that the compressors 1000, 2000, 5150 are out of
phase from each other in the same way that the above-discussed
sub-compressors 3000a, 3000b, 3000c are out of phase with each
other.
[0295] FIGS. 61-65 illustrate a compressor 4000 according to an
alternative embodiment. The compressor 4000 is generally similar to
the above-discussed compressor 2000, except that the compressor
4000 uses a pivoting gate 4050, rather than a linearly
reciprocating gate 1110. Accordingly, a redundant description of
similar or identical components is omitted. The compressor 4000
includes a main casing 4010 that defines a compression chamber 4020
(see FIGS. 61-62), a drive shaft 4030 rotationally mounted to the
casing 4010, a rotor 4040 (see FIGS. 61-62) mounted to the drive
shaft 4030 for rotation with the drive shaft 4030 relative to the
casing 4010, a gate 4050 mounted to a gate shaft 4052 for common
pivotal movement relative to the casing 4010 about a gate axis
4055, a gate-positioning system 4060, a discharge manifold 4150 in
fluid communication with an outlet 4160 into the compression
chamber 4020, and an inlet manifold 4170 in fluid communication
with an inlet 4180 of the compression chamber 4020.
[0296] As shown in FIGS. 61-62, the inlet 4180 passes through the
gate 4050. This allows for a larger inlet 4180 area as well as a
more efficient gas flow path. However, according to alternative
embodiments, the inlet 4180 may be spaced from the gate 4050
without deviating from the scope of various embodiments.
[0297] As shown in FIGS. 63-65, the gate-positioning system 4060
includes a cam 4110 mounted to the drive shaft 4030 for rotation
with the driveshaft 4030. An outer cam profile of the cam 4110
generally mimics a profile of the rotor 4040 (but may be modified
to account for pivotal-position-based changes in the way the cam
4110 drives the cam follower 4120 relative to the gate 4050), a cam
follower 4120 that abuts the cam 4110 and is mounted to the gate
shaft 4052 for common pivotal movement with the shaft 4052 and gate
4050 relative to the casing 4010 about the axis 4055 (see FIGS.
63-65), and a spring 4140 disposed between the casing 4010 and the
gate 4050 to pivotally bias the gate 4050 toward the rotor 4040. As
the rotor 4040 rotates, the gate-positioning system 4060 keeps a
seal edge 4050a of the gate proximate to the rotor 4040. The spring
4140 urges the gate 4050 toward the rotor 4040, while the cam 4110
and follower 4120 counter that force so that the seal edge 4050a
closely follows the rotor 4040 surface during operation of the
compressor 4000.
[0298] The pivoting gate 4050 helps the gate 4050 to resist the
pressure that builds up on the compressed fluid outlet 4160 side of
the gate 4050 within the compression chamber 4020. As shown in
FIGS. 61-62, the convex, semi-cylindrical surface of the gate 4050
that is exposed to high pressures in the compression volume of the
compression chamber 4020 (the right side as shown in FIGS. 61 and
62) is concentric with the gate shaft 4052 and axis 4055. As a
result, pressure loads are transferred through the gate 4050
directly to the shaft 4052 without urging the gate 4050 to pivot.
This direct force transfer through the shaft 4052 to the casing
4010 may reduce gate 4050 deflection, and reduce the forces needed
to reciprocally pivot the gate 4050 over each compression cycle of
the compressor 4000, while keeping the seal edge 4050a proximate to
the rotor 4040.
[0299] According to various embodiments, the gate 4050 and shaft
4052 may be integrally formed.
[0300] In the illustrated embodiment, a torsion spring 4140 urges
the gate 4050 toward the rotor 4040. However, any other suitable
force-imparting mechanism may alternatively be used without
deviating from the scope of the present invention (e.g., a
compression or tension spring mounted between the casing 4010 and a
lever arm attached to the gate 4050 or shaft 4052 to impart torque
on the shaft 4052 and gate 4050, a motor, magnets, etc.).
[0301] FIG. 66 illustrates a compressor 5000 according to an
alternative embodiment. The compressor 5000 is identical to the
compressor 1000, except that the compressor 5000 uses a different
type of gate support guide 5075 than the gate support guide 1075 of
the compressor 1000. A redundant description of identical
structures is omitted.
[0302] As shown in FIG. 66, the gate support guide 5075 is divided
into three parts, 5075a, 5075b, 5075c. Guide parts 5075a, 5075c
comprise gate support bushings or bearings 5080 that guide the gate
supports 5050 to permit reciprocating linear motion of the supports
5050 (in the up/down direction as illustrated in FIG. 66). The
central guide part 5075b is mounted to the casing 1010 (or
integrally formed with the casing 1010). The central guide part
5075b connects to the guide parts 5075a, 5075c via linear bearings
5090. The linear bearings 5090 permit the outer guide parts 5075a,
5075c to move toward and away from the central guide part 5075b
(i.e., along the arrows 5100 shown in FIG. 66, which extend
left/right as shown in FIG. 66). The linear bearings 5090 prevent
the outer guide parts 5075a, 5075c from moving relative to the
central guide part 5075b in a direction perpendicular to the arrows
5100 (i.e., in a direction into/out of the page as shown in FIG.
66). The linear bearings 5090 are used to correct for relative
thermal expansion of different parts of the compressor 5000 (e.g.,
between the gate support guide 5075 and the gate support cross-arm
5055), which might otherwise cause the gate support bearings 5080
to push or pull the gate supports 5050 in the direction of the
arrows 5100 and cause the supports 5050 to bind against the
bearings 5080.
[0303] According to various alternative embodiments, the linear
bearings 5090 are replaced with alternative linear movement devices
that permit the gate supports 5050 to move in the direction of the
arrows 5100. For example, thermal growth can be accounted for by
slightly undersizing the gate support 5050 relative to the linear
bearings 5080. Additionally and/or alternatively, the linear
bearings 5080 may be fitted into slotted holes in the gate casing
5075 such that the linear bearings 5080 can move axially (in the
direction of the arrows 5100) if needed due to thermal growth while
movement in a perpendicular direction (i.e., in the direction into
the page as shown in FIG. 66) is constrained or eliminated.
[0304] FIGS. 70-74 illustrate a compressor 6000 according to an
alternative embodiment. The compressor 6000 is similar to or
identical to the compressor 1000, except as explained below. A
redundant description of structures and features of the compressor
6000 that are identical or similar to structures or features of the
compressor 1000 is therefore omitted.
[0305] As shown in FIGS. 70-73, the compressor 6000 adds a casing
6010 that encloses many or all moving parts of the compressor 6000
other than the drive shaft 6020 that extends outwardly from one or
more ends of the compressor 6000.
[0306] As shown in FIG. 73, an upper portion 6030 of the casing
6010 may be integrally formed with the main casing that defines the
compression chamber 6040 of the compressor 6000. Inlet and
discharge manifolds 6050, 6060, respectively, may be integrally
formed into the upper portion 6030 of the casing 6010. The upper
portion 6030 structurally supports the hydrostatic bearing 6070 and
gate 6080, and may include reinforcing structures to stiffen the
casing and resist deflection caused by pressure from the bearing
6070 and gate 6080.
[0307] As shown in FIGS. 70 and 71, the casing 6010 also includes a
lower portion 6100 with an internal cavity that houses the springs
6110. The upper portion 6030 may bolt or otherwise removably attach
to the lower portion 6100 so that the upper portion 6030 and main
components of the compressor 6000 may be removed from the lower
portion 6100 (e.g., for maintenance or replacement). The springs
6110 may be removable as a unit along with the upper portion 6030
and main components of the compressor 6000. Alternatively, the
springs may remain with the lower portion 6100 when the upper
portion 6030 is removed.
[0308] According to various embodiments, the lower portion 6100 may
include a sump for oil from the compressor's hydraulic and
lubrication systems such that fluids reservoirs are provided within
the casing 6010.
[0309] As shown in FIG. 70, the casing 6010 also includes cam
covers 6130 that enclose and protect the cams and cam followers
(e.g., cams 1050 and followers 1060, as shown in FIG. 40). A
lubrication distribution system 6140 (e.g., an oil pump and
oil-filled reservoir) connect via conduits 6150 to the inside of
the covers 6130 to apply (e.g., spray or drip) lubricant onto the
cams and followers, and in particular the interface between the
cams and followers (shown in FIG. 39). In various embodiments, this
system may be configured to create an oil bath, wherein some
portion of the cams and cam followers may be submerged in oil for
part or all of their motion. The system may be configured to create
an optimal oil level so as to maximize lubrication provided to the
cams and cam followers while minimizing negative effects such as
oil splashing, generation of bubbles within the oil, etc. While the
system 6140 is illustrated as being on the outside of the casing
6010 in FIG. 70, the entire system 6140 and conduits 6150 may
alternatively be disposed inside the casing 6010. As shown in FIG.
72, rotational seals 6160 seal the rotational interface between the
shaft 6020 and covers 6130. Such seals 6160 may comprise mechanical
seals (e.g., rings). The seals 6160 may comprise multi-part
hydraulic seals like the seal 1500, 6200 that provide a drain and
hydrostatic over pressure to discourage working fluid that may leak
past the drive shaft into the inside of the covers 6130 from
leaking further into the ambient environment outside the covers
6130 and casing 6010.
[0310] As shown in FIG. 73, oil conduits 6170 in the upper portion
6030 may feed oil to the hydrostatic bearing 6070. The hydrostatic
bearing 6070 comprises to separate bearing pads 6070a,b (shown on
the right and left in FIG. 73) that sandwich the gate 6070
therebetween (rather than a single O or oval shaped bearing). The
two-piece bearing 6070 may facilitate grinding of the bearing 6070
and gate 6080 to reduce clearances therebetween when the bearing
6070 and gate 6080 are inserted into a matching slot in the upper
portion 6030 of the casing 6010.
[0311] As shown in FIG. 74, a gate ring mechanical/hydraulic seal
6200 surrounds the gate 6080 and seals an inside of the compression
chamber 6040 from the hydrostatic bearing 6070 and lower portion
6100 of the casing 6010. The gate ring hydraulic seal 6200 operates
in a similar manner as the seal 1500 to isolate the compression
chamber 6040 from an outer environment, except that the seal 6200
seals against the reciprocating gate 6080, rather than the rotating
drive shaft. The seal 6200 comprises, in sequential order from the
compression chamber 6040 toward the bearing 6070: a first seal
6210, a drain groove (e.g., a vent) 6220, a second seal 6230, a
hydraulic fluid groove 6240, and a third seal 6250. According to
various embodiments, the seals 6210, 6230, 6250 and grooves 6220,
6240 extend continuously around the entire perimeter of the gate
6080. The seals 6210, 6230, 6250 may each comprise single
continuous seals such as O-rings, or may comprise multi-part seals
that together form a complete perimeter around the gate 6080.
[0312] According to alternative embodiments, the seals 6210, 6230,
6250 and grooves 6220, 6240 do not extend continuously around the
gate 6080, but instead are formed by two sets of seals and grooves,
one set being disposed on the inlet side of the gate 6080 and one
set being disposed on the outlet side of the gate 6080.
[0313] As shown in FIG. 74, the drain groove (e.g., vent) 6220
fluidly connects to the inlet manifold 6050 via a fluid passageway
6280 so that working fluid that leaks from the compression chamber
6040 past the first seal 6210 is vented back into the low-pressure
inlet manifold 6050 for reinjection back into the compression
chamber 6040.
[0314] As shown in FIG. 74, the hydrostatic fluid groove 6240 is
pressurized by hydraulic fluid (or other suitable fluid) that is
pumped into the groove 6240 via a fluid passageway 6290 from a
source of pressurized fluid (e.g., hydraulic pump 1380).
[0315] As shown in FIG. 74, the seal 6200 includes a housing/body
6300 that supports the seals 6210, 6230, 6250 and grooves/vents
6220, 6240, and defines portions of the passageways 6280, 6290.
Other portions of the passageways 6280, 6290 may be defined by the
casing portion 6030 or other structures. The seal 6200 and its
components are preferably removably inserted into place within the
casing portion 6030 as a single unit. As shown in FIG. 74, the seal
6200 is inserted into a mating slot in the casing portion 6030 from
below. An additional seal ring 6310 seals the interface between the
body 6300 of the seal 6200 and the casing 6030.
[0316] The operation of the seal 6200 is described with reference
to FIG. 74. For the working fluid (e.g., natural gas being
compressed) to leak out of the compression chamber 6040 via the
opening through which the gate 6080 extends, the fluid may leak
between the seal 6210 and gate 6080. If the working fluid leaks
past the seal 6210, the fluid reaches the vent 6220, which returns
the fluid back to the low-pressure compressor inlet 6050 via the
passageway/port 6280, which is maintained at the pressure of the
inlet 6050 via its fluid communication with the inlet 6050. The
area between the second and third seals 6230, 6250 is pressurized
by hydraulic fluid fed through the passageway 6290 and groove 6240
to a pressure higher than the inlet 6050 pressure, which
discourages or prevents the working fluid from further leaking past
the seals 6230, 6250 and groove 6240. Leaked working fluid leaks
through the groove 6220 and passageway 6280 back to the intake
6050, rather than past the seals 6230, 6250 and groove 6240 because
the inlet 6050 is at a significantly lower pressure than the groove
6240. Thus, leakage of the working fluid past the seal 6200 is
reduced or preferably eliminated.
[0317] According to various alternative embodiments, additional
seals like the seals 6210, 6230, 6250 and corresponding vents like
the vents 6220, 6240 may be disposed along the leakage path between
the first of such seals and the last of such seals, which results
in a plurality of drain vents 6220 back to the inlet and/or a
plurality of pressurized vents/grooves 6240, with seals separating
the different ones of the vents/grooves 6220, 6240. According to
various embodiments, the total number of such seals along the
leakage path may comprise from 3 to 50 seals.
[0318] According to alternative embodiments, the first seal 6210
and vent 6220 may be eliminated so that the mechanical seal 6200
relies on the pressurized groove/vent 6240 to discourage leaks
across the seal 6200. According to alternative embodiments, the
third seal 6250 and vent/groove 6240 are eliminated, so that the
mechanical seal 6200 relies on the vent 6220 to discourage further
leakage past the seal 6230.
[0319] According to various embodiments, a flywheel may be added to
one or both ends of the drive shaft 6020 to reduce torsional loads
on the shaft 6020 during operation of the compressor 6000.
[0320] According to various embodiments, any of the components or
features (e.g., hydrostatic bearing 1300, mechanical seal 1500,
compression of multi-phase fluids, etc.) of any of the
above-described compressors (e.g., compressors 1000, 2000, 3000,
4000, 5000, 5150, 6000) may be used in any of the other compressors
described herein. For example, the discharge manifold 1160 may be
mounted to the outlet side 154 of the gate casing 150 of the
compressor illustrated in FIG. 28 so as to receive compressed fluid
that is expelled through outlet ports 435.
[0321] The presently preferred embodiments could be modified to
operate as an expander. Further, although descriptions have been
used to describe the top and bottom and other directions, the
orientation of the elements (e.g. the gate 600 at the bottom of the
rotor casing 400) should not be interpreted as limitations on
embodiments of the present invention.
[0322] While various of the above-described embodiments comprise a
rotary compressor that relies on a rotor that is rigidly mounted to
a drive shaft so that the rotor and drive shaft rotate together
relative to the compression chamber, various of the above-discussed
features may be used with other types of compressors (e.g., rolling
piston, screw compressor, scroll compressor, lobe, liquid ring, and
rotary vane compressors) without deviating from the scope of these
embodiments or the invention. For example, the above discussed
hydrostatic bearing arrangement 1300 can be incorporated into a
variety of other types of compressors that use moving gates/vanes
(e.g., rolling piston compressors, rotary vane compressors, etc.)
without deviating from the scope of such embodiments or the
invention.
[0323] While the foregoing written description of various
embodiments of the invention enables one of ordinary skill to make
and use what is considered presently to be the best mode thereof,
those of ordinary skill will understand and appreciate the
existence of variations, combinations, and equivalents of the
specific embodiment, method, and examples herein. The invention
should therefore not be limited by the above described embodiment,
method, and examples, but by all embodiments and methods within the
scope and spirit of the invention.
[0324] It is therefore intended that the foregoing detailed
description be regarded as illustrative rather than limiting, and
that it be understood that it is the following claims, including
all equivalents, that are intended to define the spirit and scope
of this invention. To the extent that "at least one" is used to
highlight the possibility of a plurality of elements that may
satisfy a claim element, this should not be interpreted as
requiring "a" to mean singular only. "A" or "an" element may still
be satisfied by a plurality of elements unless otherwise
stated.
* * * * *