U.S. patent application number 16/435030 was filed with the patent office on 2019-09-26 for system and method for hydrostatic bearings.
The applicant listed for this patent is Energy Recovery, Inc.. Invention is credited to David Deloyd Anderson, Chinmay Vishwas Deshpande.
Application Number | 20190293118 16/435030 |
Document ID | / |
Family ID | 54754804 |
Filed Date | 2019-09-26 |
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United States Patent
Application |
20190293118 |
Kind Code |
A1 |
Anderson; David Deloyd ; et
al. |
September 26, 2019 |
SYSTEM AND METHOD FOR HYDROSTATIC BEARINGS
Abstract
A system, includes a hydraulic transfer system configured to
exchange pressures between a first fluid and a second fluid,
wherein the first fluid has a pressure higher than the second
fluid, comprising: a sleeve; a cylindrical rotor disposed within
the sleeve in a concentric arrangement and has a first end face and
a second end face disposed opposite each other; a first end cover
having a first surface that interfaces with the first end face of
the cylindrical rotor; a second end cover having a second surface
that interfaces with the second end face of the cylindrical rotor;
and a hydrostatic bearing system configured to utilize a bearing
fluid at a pressure higher than the second fluid to resist axial
displacement, radial displacement, or both axial and radial
displacement of the cylindrical rotor.
Inventors: |
Anderson; David Deloyd;
(Castro Valley, CA) ; Deshpande; Chinmay Vishwas;
(Fremont, CA) |
|
Applicant: |
Name |
City |
State |
Country |
Type |
Energy Recovery, Inc. |
San Leandro |
CA |
US |
|
|
Family ID: |
54754804 |
Appl. No.: |
16/435030 |
Filed: |
June 7, 2019 |
Related U.S. Patent Documents
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Application
Number |
Filing Date |
Patent Number |
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14943318 |
Nov 17, 2015 |
10359075 |
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16435030 |
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62081470 |
Nov 18, 2014 |
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62088333 |
Dec 5, 2014 |
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Current U.S.
Class: |
1/1 |
Current CPC
Class: |
E21B 43/26 20130101;
F16C 32/0644 20130101; F16C 32/06 20130101; E21B 43/16 20130101;
F04F 13/00 20130101; E21B 43/267 20130101 |
International
Class: |
F16C 32/06 20060101
F16C032/06; F04F 13/00 20060101 F04F013/00; E21B 43/16 20060101
E21B043/16; E21B 43/26 20060101 E21B043/26; E21B 43/267 20060101
E21B043/267 |
Claims
1. A system, comprising: a hydraulic transfer system configured to
exchange pressures between a first fluid and a second fluid,
wherein the first fluid has a first pressure and the second fluid
has a second pressure, and wherein the first pressure is higher
than the second pressure, comprising: a cylindrical rotor
configured to rotate circumferentially about a rotational axis, the
cylindrical rotor defining a first end face and a second end face,
wherein the first end face is opposite the second end face; and a
hydrostatic bearing system configured to utilize a bearing fluid at
a pressure higher than the second pressure of the second fluid to
resist axial displacement, radial displacement, or both axial and
radial displacement of the cylindrical rotor, the hydrostatic
bearing system comprises: a first end cover having a first surface
that interfaces with the first end face of the cylindrical rotor,
the first surface defines a first groove that extends a first arc
length circumferentially about the rotational axis, wherein the
first end cover defines an aperture that extends through a
circumferential side surface of the first end cover, wherein the
aperture fluidly couples to the first groove through the first end
cover, wherein the first groove is configured to receive the
bearing fluid.
2. The system of claim 1, wherein the hydraulic transfer system
comprises a rotary isobaric pressure exchanger.
3. The system of claim 1, comprising a second end cover comprising
a second surface that interfaces with the second end face of the
cylindrical rotor.
4. The system of claim 3, wherein the second end cover defines a
second groove that extends a second arc length circumferentially
about the rotational axis and is configured to receive the bearing
fluid to provide a fluidic bearing between the second end cover and
the cylindrical rotor.
5. The system of claim 4, wherein the first surface of the first
end cover or the second surface of the second end cover comprises
at least one additional groove that extends circumferentially at
least partially about the rotational axis, the at least one
additional groove is radially offset from the first groove or the
second groove relative to the rotational axis.
6. The system of claim 1, wherein the hydrostatic bearing system is
configured to utilize the first fluid in the first groove to apply
an axial force against the first end face to block and/or reduce
contact between the cylindrical rotor and the first end cover.
7. The system of claim 1, wherein the hydrostatic bearing system
comprises a radial hydrostatic bearing system configured to resist
radial displacement of the cylindrical rotor.
8. The system of claim 7, wherein the radial hydrostatic bearing
system is configured to apply a radial force to the cylindrical
rotor to align the rotational axis of the cylindrical rotor with a
central axis of the hydraulic transfer system.
9. The system of claim 1, the first end cover defining an axial
aperture that extends into the first surface, wherein the axial
aperture and the aperture fluidly connect within the first end
cover, wherein the aperture is configured to receive the bearing
fluid and direct the bearing fluid to the axial aperture.
10. The system of claim 9, wherein an outlet of the axial aperture
is within the first groove.
11. The system of claim 1, wherein the aperture extends radially
through the first end cover.
12. A system, comprising: a hydraulic transfer system configured to
exchange pressures between a first fluid and a second fluid,
wherein the first fluid has a first pressure and the second fluid
has a second pressure, and wherein the first pressure is higher
than the second pressure, comprising: a cylindrical rotor
configured to rotate circumferentially about a rotational axis, the
cylindrical rotor defining a first end face and a second end face,
wherein the first end face is opposite the second end face; and a
hydrostatic bearing system configured to utilize a bearing fluid at
a pressure higher than the second pressure of the second fluid to
resist axial displacement, radial displacement, or both axial and
radial displacement of the cylindrical rotor, the hydrostatic
bearing system comprises: a first end cover comprising a first
surface that interfaces with the first end face of the cylindrical
rotor and a first circumferential side surface, the first end cover
defines a first plurality of axial apertures that extend into the
first surface and a first plurality of radial apertures that extend
into the first circumferential side surface, wherein the first
plurality of axial apertures fluidly connect to the first plurality
of radial apertures within the first end cover, wherein the first
plurality of radial apertures are configured to receive the bearing
fluid and direct the bearing fluid to the first plurality of axial
apertures.
13. The system of claim 12, wherein the first plurality of axial
apertures are circumferentially spaced about the first surface.
14. The system of claim 12, wherein the first plurality of axial
apertures are equally spaced from a central axis of the first end
cover.
15. The system of claim 12, wherein the first plurality of axial
apertures fluidly connects to a respective radial aperture of the
first plurality of radial apertures.
16. The system of claim 12, comprising a second end cover
comprising a second surface that interfaces with the second end
face of the cylindrical rotor, the second end cover defines a
second plurality of axial apertures that extend into the second
surface and a second plurality of radial apertures that extend into
a second circumferential side surface, wherein the second plurality
of axial apertures fluidly connect to the second plurality of
radial apertures within the second end cover, wherein the second
plurality of radial apertures are configured to receive the bearing
fluid and direct the bearing fluid to the second plurality of axial
apertures.
17. A rotary isobaric pressure exchanger configured to exchange
pressures between a first fluid and a second fluid, wherein the
first fluid has a first pressure and the second fluid has a second
pressure, and wherein the first pressure is higher than the second
pressure, the rotary isobaric pressure exchanger comprising: a
cylindrical rotor configured to rotate circumferentially about a
rotational axis, the cylindrical rotor defining a first end face
and a second end face, wherein the first end face is opposite the
second end face; a hydrostatic bearing system configured to utilize
a bearing fluid at a pressure higher than the second pressure of
the second fluid to resist axial displacement, radial displacement,
or both axial and radial displacement of the cylindrical rotor, the
hydrostatic bearing system comprises: a first end cover comprising
a first surface that interfaces with the first end face of the
cylindrical rotor and a first circumferential side surface, the
first end cover defines a first plurality of axial apertures that
extend into the first surface and a first plurality of radial
apertures that extend into the first circumferential side surface,
wherein the first plurality of axial apertures fluidly connect to
the first plurality of radial apertures within the first end cover,
wherein the first plurality of radial apertures are configured to
receive the bearing fluid and direct the bearing fluid to the first
plurality of axial apertures; and a second end cover comprising a
second surface that interfaces with the second end face of the
cylindrical rotor, the second end cover defines a second plurality
of axial apertures that extend into the second surface and a second
plurality of radial apertures that extend into a second
circumferential side surface, wherein the second plurality of axial
apertures fluidly connect to the second plurality of radial
apertures within the second end cover, wherein the second plurality
of radial apertures are configured to receive the bearing fluid and
direct the bearing fluid to the second plurality of axial
apertures.
18. The rotary isobaric pressure exchanger of claim 17, wherein the
first surface of the first end cover defines a first plurality of
grooves wherein an outlet of each aperture of the first plurality
of axial apertures is within a respective groove of the first
plurality of grooves.
19. The rotary isobaric pressure exchanger of claim 17, wherein the
second surface of the second end cover defines a second plurality
of grooves wherein an outlet of each aperture of the second
plurality of axial apertures is within a respective groove of the
second plurality of grooves.
20. The rotary isobaric pressure exchanger of claim 17, wherein the
hydrostatic bearing system comprises a radial hydrostatic bearing
system configured to resist radial displacement of the cylindrical
rotor.
Description
CROSS-REFERENCE TO RELATED APPLICATION
[0001] This application claims priority to and benefit of U.S.
patent application Ser. No. 14/943,318, entitled "SYSTEM AND METHOD
FOR HYDROSTATIC BEARINGS," filed on Nov. 17, 2015, which claims
priority to and benefit of U.S. Provisional Patent Application No.
62/081,470, entitled "SYSTEMS AND METHODS FOR AN AXIAL HYDROSTATIC
BEARING," filed on Nov. 18, 2014, and U.S. Provisional Patent
Application No. 62/088,333, entitled "MULTI-POCKET HYDROSTATIC
BEARINGS," filed on Dec. 5, 2014, which are hereby incorporated by
reference in their entirety for all purposes.
BACKGROUND
[0002] This section is intended to introduce the reader to various
aspects of art that may be related to various aspects of the
present invention, which are described and/or claimed below. This
discussion is believed to be helpful in providing the reader with
background information to facilitate a better understanding of the
various aspects of the present invention. Accordingly, it should be
understood that these statements are to be read in this light, and
not as admissions of prior art.
[0003] The subject matter disclosed herein relates to rotating
equipment, and, more particularly, to systems and methods for an
axial bearing system for use with rotating equipment.
[0004] Rotating equipment, such as pumps, may handle a variety of
fluids. In certain applications, axial pressure imbalances (i.e.
the difference in average pressure between the two axial faces) may
exert a substantial net force on rotating components of the
rotating equipment. Axial forces may also arise due to the weight
of the rotating components. Various bearings may be used to
facilitate the rotation of the rotating components of the
equipment. However, in situations that require a high pressure
and/or a challenging environment, rotating equipment may require
additional or increased bearing capacity and functionality. For
example, in some situations, rotating equipment with insufficient
bearing capacity may result in axial contact between rotating
components and stationary components resulting in stalling, wear,
stress, and may reduce the life of the equipment and result in a
loss of efficiency. Accordingly, it may be beneficial to provide
rotating equipment with features that provide additional bearing
capacity, such as, for example, additional load bearing capacity or
additional stiffness.
BRIEF DESCRIPTION OF THE DRAWINGS
[0005] Various features, aspects, and advantages of the present
invention will become better understood when the following detailed
description is read with reference to the accompanying figures in
which like characters represent like parts throughout the figures,
wherein:
[0006] FIG. 1 is a schematic diagram of an embodiment of a
hydraulic energy transfer system having a hydrostatic bearing
system;
[0007] FIG. 2 is a cross-sectional diagram of an embodiment of the
hydraulic energy transfer system of FIG. 1;
[0008] FIG. 3 is a schematic diagram of an embodiment of a frac
system with a hydraulic energy transfer system;
[0009] FIG. 4 is an exploded perspective view of an embodiment of
the hydraulic energy transfer system of FIG. 1, illustrated as a
rotary isobaric pressure exchanger (IPX) system;
[0010] FIG. 5 is an exploded perspective view of an embodiment of a
rotary IPX in a first operating position;
[0011] FIG. 6 is an exploded perspective view of an embodiment of a
rotary IPX in a second operating position;
[0012] FIG. 7 is an exploded perspective view of an embodiment of a
rotary IPX in a third operating position;
[0013] FIG. 8 is an exploded perspective view of an embodiment of a
rotary IPX in a fourth operating position;
[0014] FIG. 9 is a schematic cross-sectional view of a rotary
IPX;
[0015] FIG. 10 is a schematic cross-sectional view of a rotary IPX
in an unbalanced position;
[0016] FIG. 11 is a schematic cross-sectional view of a rotary IPX
having radial and axial bearings;
[0017] FIG. 12 is a schematic diagram of an embodiment of the
hydraulic energy transfer system of FIG. 1, illustrating the
hydrostatic bearing system with an endcover notch disposed within
the plenum region;
[0018] FIG. 13 is a schematic diagram of an embodiment of the
hydraulic energy transfer system of FIG. 1, illustrating the
hydrostatic bearing system with a rotor notch disposed within the
plenum region;
[0019] FIG. 14 is a perspective view of an embodiment of an end
cover of a rotary IPX having axial bearings;
[0020] FIG. 15 is a cross-sectional view of an embodiment of the
end cover taken along line 15-15 of the rotary IPX of FIG. 14;
[0021] FIG. 16 is a sectional view of an embodiment of the axial
bearing of FIG. 14 within line 16-16;
[0022] FIG. 17 is a perspective view of a sleeve of the rotary IPX
of FIG. 3 having radial bearings;
[0023] FIG. 18 is a partial cross-sectional view of an embodiment
of the radial bearing of FIG. 17 within the line 18-18;
[0024] FIG. 19 is a partial cross-sectional view of an embodiment
of the radial bearing of FIG. 17 within the line 19-19;
[0025] FIG. 20 is a cross-sectional diagram of an embodiment of an
endcover of the hydraulic energy transfer system of FIG. 1,
illustrating a sink channel and a low pressure sink;
[0026] FIG. 21 is a cross-sectional diagram of an embodiment of the
endcover of FIG. 20, illustrating a first sink channel, a second
sink channel, and a partial low pressure sink loop; and
[0027] FIG. 22 is a cross-sectional diagram of an embodiment of the
endcover of FIG. 20, illustrating the sink channel, a partial low
pressure sink, and one or more low pressure sink endpoints.
DETAILED DESCRIPTION OF SPECIFIC EMBODIMENTS
[0028] One or more specific embodiments of the present invention
will be described below. These described embodiments are only
exemplary of the present invention. Additionally, in an effort to
provide a concise description of these exemplary embodiments, all
features of an actual implementation may not be described in the
specification. It should be appreciated that in the development of
any such actual implementation, as in any engineering or design
project, numerous implementation-specific decisions must be made to
achieve the developers' specific goals, such as compliance with
system-related and business-related constraints, which may vary
from one implementation to another. Moreover, it should be
appreciated that such a development effort might be complex and
time consuming, but would nevertheless be a routine undertaking of
design, fabrication, and manufacture for those of ordinary skill
having the benefit of this disclosure.
[0029] As discussed in detail below, the embodiments disclosed
herein generally relate to systems and methods for rotating systems
that may be utilized in various industrial applications. The
rotating systems disclosed herein may include a hydrostatic bearing
system configured to provide an additional bearing capacity, so
that the rotating system provides sufficient load capacity for
supporting certain rotating equipment, such as the rotor. Indeed,
in certain industrial situations involving high pressures or other
challenging applications, the bearing system of the rotating system
may have insufficient load capacities or functionalities to support
rotating equipment, such as the rotor. Such situations may result
in a stalled rotor and/or contact/friction between portions of the
bearing system, thereby resulting in a loss of efficiency, wear,
stress, and/or a reduced life of the rotating equipment.
Accordingly, the embodiments disclosed herein may provide rotating
system having a hydrostatic bearing system configured to handle
additional bearing or load capacities, which may, for example,
provide additional axial load capacities and a greater stiffness
for supporting rotating equipment, such as the rotor. Particularly,
the hydrostatic bearing system may be utilized within industrial
applications having higher pressures and/or more challenging
applications, such as, for example, in isobaric pressure
exchangers.
[0030] In certain embodiments, the rotating system may include a
hydraulic energy transfer system that is configured to handle a
variety of fluids. Specifically, the hydraulic energy transfer
system may transfer work and/or pressure between first and second
fluids via a hydrostatic bearing system that may be used to
facilitate the rotation of rotating components of the equipment.
Generally, hydrostatic bearing systems within the hydraulic energy
transfer system may operate with a source of fluid (e.g., high
pressure bearing fluid) that is introduced between a rotor and
endcovers (e.g., support for the rotor). The high pressure of the
fluid source may be configured to support the rotor on a fluid film
and may be configured to facilitate the rotation of the rotating
components. Particularly, when the rotor moves away from the
endcovers, an axial clearance region between the rotor and the
endcover may increase. The increase in the axial clearance region
allows the high pressure fluid to escape, thereby decreasing the
pressure acting on the rotor. Likewise, when the axial clearance
region is small between the rotor and the endcover, high pressure
fluid builds within the axial bearing region.
[0031] The hydrostatic bearing system may include one or more axial
clearance openings within the axial bearing region (e.g., the
interface between the rotor and the end cover of the hydraulic
energy transfer system) that help provide additional bearing
capacities and/or axial load capacities for the hydraulic energy
transfer system. The axial clearance opening may provide a greater
hydrostatic bearing restoring force that results in a stiffer
hydrostatic bearing system, which may increase the overall axial
bearing capacity of the hydraulic energy transfer system. In
certain embodiments, the hydrostatic bearing system may include a
low pressure sink connected to the low pressure region of the
hydraulic energy transfer system via one or more sink channels.
[0032] The hydraulic energy transfer system may include a hydraulic
turbocharger or a hydraulic pressure exchange system, such as a
rotating isobaric pressure exchanger (IPX). In some embodiments,
the pressures of the volumes of first and second fluids may not
completely equalize. Thus, in certain embodiments, the IPX may
operate isobarically, or the IPX may operate substantially
isobarically (e.g., wherein the pressures equalize within
approximately +/-1, 2, 3, 4, 5, 6, 7, 8, 9, or 10 percent of each
other). In certain embodiments, a first pressure of a first fluid
(e.g., pressure exchange fluid, motive fluid, etc.) may be greater
than a second pressure of a second fluid (e.g., corrosive fluid).
For example, the first pressure may be between approximately 5,000
kPa to 25,000 kPa, 20,000 kPa to 50,000 kPa, 40,000 kPa to 75,000
kPa, 75,000 kPa to 100,000 kPa or greater than the second pressure.
Thus, the IPX may be used to transfer pressure from a first fluid
(e.g., pressure exchange fluid, motive fluid, etc.) at a higher
pressure to a second fluid (e.g., corrosive fluid) at a lower
pressure. In particular, during operation, the hydraulic energy
transfer system may help block or limit contact between the
corrosive fluid and other equipment within the industrial
applications (e.g., pumps). By blocking or limiting contact between
pumps and the corrosive fluids, the hydraulic energy transfer
system increases the life/performance while reducing abrasion/wear
of various high pressure pumps within various industrial
applications as described in detail below.
[0033] FIG. 1 is a schematic diagram of an embodiment of a
hydraulic energy transfer system 10. In particular, in the
illustrated embodiment, the hydraulic energy transfer system 10
(e.g., a hydraulic turbocharger or IPX) may be configured to
transfer energy from a first fluid to a second fluid. Furthermore,
in certain embodiments, the hydraulic energy transfer system 10 may
include a hydrostatic bearing system 12 configured with features
that help provide the hydraulic energy transfer system 10 with
additional bearing capacities and/or additional axial load
capacities.
[0034] In certain embodiments, the hydraulic energy transfer system
10 may be configured with a rotary IPX 20 configured to receive a
first fluid and a second fluid. It should be noted that reference
to various directions (e.g., axial direction 32, radial direction
142, and circumferential direction 148) may be referred to in the
following discussion. In certain embodiments, a high pressure pump
may be configured to pump the first fluid to the hydraulic energy
transfer system 10 at a high pressure. For example, as illustrated,
the first fluid may be provided as a high pressure first fluid
inlet 14 to the hydraulic energy transfer system 10. Further, in
certain embodiments, a low pressure pump may be configured to pump
the second fluid to the hydraulic energy transfer system 10 at a
low pressure. For example, as illustrated, the second fluid may be
provided as a low pressure second fluid inlet 16 to the hydraulic
energy transfer system 10. During operation, the hydraulic energy
transfer system 10 may be configured to transfer pressures between
the first fluid and the second fluid.
[0035] As used herein, the isobaric pressure exchanger (IPX) 20 may
be generally defined as a device that transfers fluid pressure
between a high pressure inlet stream and a low pressure inlet
stream at efficiencies in excess of approximately 50%, 60%, 70%, or
80% without utilizing centrifugal technology. In this context, high
pressure refers to pressures greater than the low pressure. The low
pressure inlet stream of the IPX 20 may be pressurized and exit the
IPX at high pressure (e.g., at a pressure greater than that of the
low pressure inlet stream), and the high pressure inlet stream may
be depressurized and exit the IPX 20 at low pressure (e.g., at a
pressure less than that of the high pressure inlet stream).
Additionally, the IPX 20 may operate with the high pressure fluid
directly applying a force to pressurize the low pressure fluid,
with or without a fluid separator between the fluids. Examples of
fluid separators that may be used with the IPX include, but are not
limited to, pistons, bladders, diaphragms and the like. In certain
embodiments, isobaric pressure exchangers 20 may be rotary devices.
Rotary isobaric pressure exchangers (IPXs) 20, such as those
manufactured by Energy Recovery, Inc. of San Leandro, Calif., may
not have any separate valves, since the effective valving action is
accomplished internal to the device via the relative motion of a
rotor with respect to end covers, as described in detail below with
respect to FIGS. 5-8. Rotary IPXs 20 may be designed to operate
with internal pistons to isolate fluids and transfer pressure with
relatively little mixing of the inlet fluid streams. Reciprocating
IPXs may include a piston moving back and forth in a cylinder for
transferring pressure between the fluid streams. Any IPX or
plurality of IPXs 20 may be used in the disclosed embodiments, such
as, but not limited to, rotary IPXs, reciprocating IPXs, or any
combination thereof. In addition, the IPX 20 may be disposed on a
skid separate from the other components of a fluid handling system,
which may be desirable in situations in which the IPX 20 is added
to an existing fluid handling system.
[0036] As noted above, in certain embodiments, the hydraulic energy
transfer system 10 may include the hydrostatic bearing system 12
configured to help facilitate the rotation of the rotating
components within the system, such as the rotor 44. Generally, a
high pressure bearing fluid may be introduced in proximity to an
axial midplane 18 of the rotor 44, about the circumference where it
may circulate toward the axial faces where it facilitates radial
and axial load bearing. In particular, the plenum region 22
includes an inner wall of a sleeve 24 of the IPX system 20, an
outer wall of the rotor 44 disposed within the IPX system 20, and a
gap 26 between the sleeve 24 and the rotor 44. In certain
embodiments, the plenum region 22 includes an axial bearing region
28, which may be the gap 26 between the rotor 44 and the endcovers
30. The high pressure bearing fluid introduced into the plenum
region 22 may be configured to support the rotor 44 on a fluid film
and may be configured to facilitate the rotation of the rotor 44.
Particularly, when the rotor 44 moves axially 32 away from the
endcover 30 due to forces caused by the high pressure bearing
fluid, an axial clearance region 34 between the rotor 44 and the
endcover 30 may increase. The increase in the axial clearance
region 34 allows the high pressure fluid to escape, thereby
decreasing a net force acting on the rotor 44 and reducing the
axial clearance region 34 to the lesser amount. Likewise, when the
axial clearance region 34 is small between the rotor 44 and the
endcover 30, high pressure fluid builds within the axial bearing
region 28 resulting in a restoring force that increases the axial
clearance region 34. In this manner, components of the hydrostatic
bearing may work in tandem to create a stiff rotor that resists
axial displacement and facilities the steady rotation of the rotor
44. In certain embodiments, the hydrostatic bearing system 12 may
include additional features that provide the IPX system 20 with
additional load bearing capacities, such as an additional axial
load capacity, as further described with respect to FIGS. 12 and
20-22.
[0037] FIG. 2 is a cross-sectional diagram taken along the line 2-2
of the hydraulic energy transfer system 10 of FIG. 1. Specifically,
the illustrated embodiment depicts a high pressure region 36 and a
low pressure region 38 of the hydraulic energy transfer system 10
along the axial clearance region 34 with respect to the axial
surface of the endcover 30. For example, the high pressure region
36 along the axial clearance region 34 may be proximate to the high
pressure first fluid inlet 14 and a high pressure second fluid
outlet 40. Further, the low pressure region 38 along the axial
clearance region 34 may be proximate to a low pressure first fluid
outlet 42 and the low pressure second fluid inlet 16. Accordingly,
the low pressure region 38 of the hydrostatic bearing system 12 may
be, in certain embodiments, concentrated proximate to the low
pressure first fluid outlet 42 and the low pressure second fluid
inlet 16 along the axial clearance region 34. In some embodiments,
the hydraulic energy transfer system 10 may be used with a frac
system to capture and recycle otherwise wasted pressure energy in
fluid flows, as discussed further with respect to FIG. 3.
[0038] FIG. 3 is a schematic diagram of an embodiment of a frac
system 46 (e.g., fluid handling system) that may be used with the
hydraulic energy transfer system 10. In operation, the frac system
46 enables well completion operations to increase the release of
oil and gas in rock formations. The frac system 46 may include one
or more first fluid pumps 48 and one or more second fluid pumps 50
coupled to a hydraulic energy transfer system 10. As described
above, the hydraulic energy system 10 may include a hydraulic
turbocharger, rotary IPX, reciprocating IPX, or any combination
thereof. In addition, the hydraulic energy transfer system 10 may
be disposed on a skid separate from the other components of a frac
system 46, which may be desirable in situations in which the
hydraulic energy transfer system 10 is added to an existing frac
system 46. In operation, the hydraulic energy transfer system 10
transfers pressures without any substantial mixing between a first
fluid (e.g., proppant free fluid) pumped by the first fluid pumps
48 and a second fluid (e.g., proppant containing fluid or frac
fluid) pumped by the second fluid pumps 50. In this manner, the
hydraulic energy transfer system 10 blocks or limits wear on the
first fluid pumps 48 (e.g., high-pressure pumps), while enabling
the frac system 46 to pump a high-pressure frac fluid into the well
52 to release oil and gas. In addition, because the hydraulic
energy transfer system 10 is configured to be exposed to the first
and second fluids, the hydraulic energy transfer system 10 may be
made from materials resistant to corrosive and abrasive substances
in either the first and second fluids. For example, the hydraulic
energy transfer system 10 may be made out of ceramics (e.g.,
alumina, cermets, such as carbide, oxide, nitride, or boride hard
phases) within a metal matrix (e.g., Co, Cr or Ni or any
combination thereof) such as tungsten carbide in a matrix of CoCr,
Ni, NiCr or Co.
[0039] Returning now to the hydraulic energy transfer system 10 of
FIG. 1, the IPX system 20 may be further understood in the
discussion with respect to FIGS. 4-8. FIG. 4 is an exploded view of
an embodiment of the rotary IPX 20. In the illustrated embodiment,
the rotary IPX 20 may include a generally cylindrical body portion
54 that includes a housing 56 and the rotor 44. The rotary IPX 20
may also include two end structures 58 and 60 that include
manifolds 62 and 64, respectively. Manifold 62 includes inlet and
outlet ports 66 and 68 and manifold 64 includes inlet and outlet
ports 70 and 72. For example, inlet port 66 may receive a high
pressure first fluid and the outlet port 68 may be used to route a
low pressure first fluid away from the IPX 20. Similarly, inlet
port 70 may receive a low pressure second fluid and the outlet port
72 may be used to route a high pressure second fluid away from the
IPX 20. The end structures 58 and 60 include generally flat end
plates 31, 33 (e.g., endcovers 30), respectively, disposed within
the manifolds 62 and 64, respectively, and adapted for fluid
sealing contact with the rotor 44. The rotor 44 may be cylindrical
and disposed in the housing 56, and is arranged for rotation about
a longitudinal axis 74 of the rotor 44. The rotor 44 may have a
plurality of channels 76 extending substantially longitudinally
through the rotor 44 with openings 78 and 80 at each end arranged
symmetrically about the longitudinal axis 74. The openings 78 and
80 of the rotor 44 are arranged for hydraulic communication with
the end plates 62 and 64, and inlet and outlet apertures 82 and 84,
and 86 and 88, in such a manner that during rotation they
alternately hydraulically expose fluid at high pressure and fluid
at low pressure to the respective manifolds 62 and 64. The inlet
and outlet ports 66, 68, 70, and 72, of the manifolds 62 and 64
form at least one pair of ports for high pressure fluid in one end
element 58 or 60, and at least one pair of ports for low pressure
fluid in the opposite end element, 58 or 60. The end plates 62 and
64, and inlet and outlet apertures 82 and 84, and 86 and 88 are
designed with perpendicular flow cross sections in the form of arcs
or segments of a circle.
[0040] With respect to the IPX 20, the plant operator has control
over the extent of mixing between the first and second fluids,
which may be used to improve the operability of the fluid handling
system. For example, varying the proportions of the first and
second fluids entering the IPX 20 allows the plant operator to
control the amount of fluid mixing within the fluid handling
system. In certain embodiments, the proportion of the motive fluid
may be varied with respect to the corrosive fluid to control the
amount of mixing within the fluid handling system, as further
described with respect to FIG. 12. Three characteristics of the IPX
20 that affect mixing are: (1) the aspect ratio of the rotor
channels 76, (2) the short duration of exposure between the first
and second fluids, and (3) the creation or presence of a fluid
barrier (e.g., an interface) between the first and second fluids
within the rotor channels 76. First, the rotor channels 76 are
generally long and narrow, which stabilizes the flow within the IPX
20. In addition, the first and second fluids may move through the
channels 76 in a plug flow regime with very little axial mixing.
Second, in certain embodiments, at a rotor speed of approximately
1200 RPM, the time of contact between the first and second fluids
may be less than approximately 0.15 seconds, 0.10 seconds, or 0.05
seconds, which again limits mixing of the streams. Third, a small
portion of the rotor channel 76 is used for the exchange of
pressure between the first and second fluids. Therefore, a volume
of fluid remains in the channel 76 as a barrier between the first
and second fluids. All these mechanisms may limit mixing within the
IPX 20.
[0041] FIGS. 5-8 are exploded views of an embodiment of the rotary
IPX 20 illustrating the sequence of positions of a single channel
76 in the rotor 44 as the channel 76 rotates through a complete
cycle, and are useful to an understanding of the rotary IPX 20. It
is noted that FIGS. 5-8 are simplifications of the rotary IPX 20
showing one channel 76 and the channel 76 is shown as having a
circular cross-sectional shape. In other embodiments, the rotary
IPX 20 may include a plurality of channels 76 (e.g., 2 to 100) with
different cross-sectional shapes. Thus, FIGS. 5-8 are
simplifications for purposes of illustration, and other embodiments
of the rotary IPX 20 may have configurations different from that
shown in FIGS. 5-8. As described in detail below, the rotary IPX 20
facilitates a hydraulic exchange of pressure between two liquids by
putting them in momentary contact within a rotating chamber. In
certain embodiments, this exchange happens at a high speed that
results in very high efficiency with very little mixing of the
liquids.
[0042] In FIG. 5, the channel opening 78 is in hydraulic
communication with aperture 84 in endplate 31 and therefore with
the manifold 62 at a first rotational position of the rotor 44 and
opposite channel opening 80 is in hydraulic communication with the
aperture 88 in endplate 33, and thus, in hydraulic communication
with manifold 64. As discussed below, the rotor 44 rotates in the
clockwise direction indicated by arrow 90. As shown in FIG. 5, low
pressure second fluid 92 passes through end plate 31 and enters the
channel 76, where it pushes first fluid 94 out of the channel 76
and through end plate 31, thus exiting the rotary IPX 20. In
certain embodiments, the first and second fluids 92 and 94 contact
one another at an interface 96 where minimal mixing of the liquids
occurs because of the short duration of contact. In certain
embodiments, the interface 96 may be a direct contact interface
because the second fluid 92 directly contacts the first fluid 94.
In other embodiments, the interface 96 may include a dynamic
barrier that is utilized to separate the first fluid and the second
fluid. In other embodiments, asymmetrical flow of the first and
second fluids may result in a certain amount of mixing between the
first and second fluids.
[0043] In FIG. 6, the channel 76 has rotated clockwise through an
arc of approximately 90 degrees, and the outlet 80 is now blocked
off between apertures 86 and 88 of end plate 33, and outlet 78 of
the channel 76 is located between the apertures 82 and 84 of end
plate 31 and, thus, blocked off from hydraulic communication with
the manifold 62 of end structure 58. Thus, the low pressure second
fluid 92 is contained within the channel 76.
[0044] In FIG. 7, the channel 76 has rotated through approximately
180 degrees of arc from the position shown in FIG. 5. Opening 80 is
in hydraulic communication with aperture 86 in end plate 33 and in
hydraulic communication with manifold 64, and the opening 78 of the
channel 76 is in hydraulic communication with aperture 82 of end
plate 31 and with manifold 62 of end structure 58. The liquid in
channel 76, which was at the pressure of manifold 64 of end
structure 60, transfers this pressure to end structure 58 through
outlet 78 and aperture 82, and comes to the pressure of manifold 62
of end structure 58. Thus, high pressure first fluid 94 pressurizes
and displaces the second fluid 92.
[0045] In FIG. 8, the channel 76 has rotated through approximately
270 degrees of arc from the position shown in FIG. 3, and the
openings 70 and 72 of channel 68 are between apertures 82 and 84 of
end plate 31, and between apertures 86 and 88 of end plate 33.
Thus, the high pressure first fluid 94 is contained within the
channel 76. When the channel 76 rotates through approximately 360
degrees of arc from the position shown in FIG. 5, the second fluid
92 displaces the first fluid 94, restarting the cycle.
[0046] The rotary IPX system 20 may be further understood with
respect to the discussion of FIGS. 9-11. FIGS. 9-11 depict various
cross sectional views of the IPX 20. FIG. 9 is a schematic
cross-sectional view of an embodiment of the rotary IPX 20. It will
be appreciated that FIG. 9 is a simplified view of the rotatory IPX
20 and certain details have been omitted for clarity. As described
above, the rotary IPX 20 includes the housing 56 containing the
sleeve 24, the rotor 44, the end covers 31, 33, and a seal 98 among
other components. As illustrated, the seal 98 may be disposed
between the housing 56 and the end covers 31, 33 to substantially
block the flow of the first fluid 94 from exiting the housing 56.
However, in the illustrated embodiment, the seal 98 is not
positioned about the end cover 31. As discussed above, the HP first
fluid 94 may enter the rotary IPX 20 through the inlet 66 and the
aperture 82 to drive the LP second fluid 92 out of the channel 76.
In certain embodiments, the HP first fluid 94 entering the rotary
IPX 20 may transfer a first force 100 to a face of the rotor 20.
Additionally, an axial force may also be generated by the LP second
fluid 92 entering the rotary IPX 20. For example, as the rotor 20
rotates within the sleeve 24, the HP first fluid 94 may contact
solid portions of the rotor 44 between the channels 76.
Accordingly, the first force 100 may drive the rotor 44 toward the
end cover 186. Moreover, because the first force 100 acts on
substantially a top half of the rotor 44, the first force 100 may
shift the rotor toward the sleeve 24. That is, the rotor 44 may
tilt relative to the axis 74. As a result, the likelihood that the
rotor 44 is unbalanced (e.g., an axis of the rotor 44 is not
substantially coaxial with the axis 74) during operation increases,
thereby potentially increasing the likelihood for friction and
wear.
[0047] FIG. 10 is a schematic cross-sectional view of an embodiment
of the rotor 44 in an unbalanced position as a result of the first
force 100. In an unbalanced position, a rotor axis 116 may be
positioned at an angle 102, relative to the axis 74; instead of
substantially aligned (e.g., coaxial). Accordingly, a first end 104
of the rotor 44 is positioned at a closer distance 108 to the
sleeve 24 than a second end 106 of the rotor 44 at a distance 110,
relative to the sleeve 24. Moreover, the first force 100 may drive
the first end 222 of the rotor 166 a distance 112 from the end
cover 31, reducing a distance 114 between the second end 106 and
the end cover 33. Accordingly, the likelihood of the rotor 44
contacting the end cover 33 and the sleeve 24 may increase because
of the first force 100.
[0048] FIG. 11 is a schematic cross-sectional view of rotary IPX 44
having axial and radial bearings 118, 120 (e.g., fluid bearings,
hydraulic pressure bearings) configured to maintain the rotary IPX
20 in a balanced position. As used herein, fluid bearings may
include fluid dynamic bearings, hydrodynamic bearings, hydrostatic
bearings, and the like. Axial bearings 118 are disposed on a face
122 of the end cover 33. In certain embodiments, the axial bearings
118 are configured to apply a force 124 (e.g., a hydraulic force or
pressure) to the rotor 44 to drive the rotor 44 away from the end
cover 33. That is, the axial bearings 118 are configured to drive
the rotor 44 toward the end cover 31, thereby maintaining a
clearance 126 between the rotor 44 and end cover 33 and to block
contact between the rotor 44 and the end cover 33. As will be
described in detail below, in certain embodiments, the axial
bearings 118 may be positioned at a variety of locations on the
face 122 of the end cover 33 to focus the force 124 at desired
locations on the rotor 44. The axial bearings 118 may also balance
the rotor 44 (e.g., align the rotor axis 116 with the axis 74 of
the rotary IPX 20).
[0049] In the illustrated embodiments, the axial bearings 118 are
fed the first fluid 94 via flow passages 128 on an outer
circumference of the end cover 31, 33. The flow passages 128
receive the HP first fluid 94 disposed outside of the sleeve 24.
For example, in certain embodiments, HP first fluid 94 may enter an
outer channel 130 via a gap 132 (e.g., axial gap) between the end
cover 31 and the rotor 44 or a gap 134 (e.g., radial gap) between
the end cap 31 and the housing 56. The outer channel 130 (e.g.,
axial and/or annular channel) may direct the HP first fluid 94
toward the flow passages 128, thereby enabling the axial bearings
118 to utilize pressure of the HP first fluid 94 to apply a force
on the rotor 44 via outlets distributed on an axial face of the end
covers 31, 33. The outlets enable the axial bearings 118 to apply
the force and/or pressure of the first fluid 94 to the rotor 44. It
should be noted that, in other embodiments, the outer channel 130
may receive the HP first fluid 94 via passages in the end cover 31,
a flow inlet in the housing 56, or the like.
[0050] In some embodiments, the sleeve 24 may also include radial
bearings 120 (e.g., fluid bearings). The radial bearings 120 extend
through the sleeve 24 and are configured to use the pressure of the
first fluid 94 to drive the rotor 44 away from the sleeve 24.
Similar to the axial bearings 118, the HP first fluid 94 feeds the
radial bearings 120 via the outer channel 130 to utilize the
pressure of the HP first fluid 94 to apply a force 136 on the rotor
44. As a result, the radial bearing 120 may align the rotor axis
116 with the axis 74 to block tilting of the rotor 44. Moreover,
the first fluid 94 may clean out or purge the axial and radial
bearings 118, 120 at the interface between rotating and stationary
parts. For example, in the illustrated embodiment, the first fluid
94 is clean or substantially debris-free. As a result, the first
fluid 94 may drive particles positioned at the interfaces away,
thereby helping to reduce any abrasion, wear, friction, or the like
along the interface between the rotating and stationary parts.
[0051] FIG. 12 is a schematic diagram of an embodiment of the
hydraulic energy transfer system 10 of FIG. 1, illustrating the
hydrostatic bearing system 12 with an axial clearance opening 126
disposed within the axial bearing region 28. Specifically, in the
illustrated embodiment, the axial clearance opening 126 may be
formed by forming an endcover notch 136 along a portion of the
endcover 31, 33, such as along the interface 138 between the rotor
44 and the endcover 31. For example, in certain embodiments, the
endcover notch 136 may be disposed on a first surface 150 of the
endcover 31, 33 which may be facing and parallel to a second
surface 152 of the rotor 44. In certain embodiments, the axial
clearance opening 126 and the endcover notch 136 may be configured
to spatially increase the axial bearing region 28 of the hydraulic
energy transfer system 10, thereby resulting in an increased
restoring force that acts on the rotor 44, as further described in
detail below. Further, the greater restoring force acting on the
rotor 44 may result in a stiffer hydrostatic bearing system, which
may help provide a greater axial load capacity for the IPX system
20, as further described below.
[0052] In certain embodiments, the interface between the rotor 44
and the endcover 31 may be substantially flat and the first surface
150 of the endcover and a second surface 154 of the endcover may be
parallel to each other. In the illustrated embodiment, the axial
clearance opening 136 may be formed by forming an endcover notch
136 along a portion of the first surface 150 of the endcover. The
axial clearance opening 126 and the endcover notch 136 may be
configured to spatially increase the axial clearance region 34 of
the IPX system 20. For example, the increased axial clearance
region 34 may allow a greater amount of the high pressure bearing
fluid to build up, thereby increasing the amount of restoring force
of the hydrostatic bearing system 12. As noted above, the restoring
force may be configured to increase the axial clearance region 34
or gap between the rotor 44 and the endcover 31, 33. It should be
noted that increasing the axial clearance region 34 additionally
increases the amount of high pressure close to the high pressure
region of the IPX system 20. In this manner, the pressure in the
axial clearance region 34 may have more efficacy in providing a
restoring force to the rotor 44, which results in a stiffer
hydrostatic bearing system 12 that has a higher load capacity.
[0053] In certain embodiments, the position and size of the
endcover notch 136 may be optimized based on the amount of load
capacity or stiffness desired from the hydrostatic bearing system
12. For example, in order to maximize the stiffness of the
hydrostatic bearing system 12, the flow resistance from the high
pressure region 36 to the axial clearance region 34 should be
approximately the same as the flow resistance from the axial
clearance region 34 to the low pressure region 38. Further,
adjusting the spatial size of the endcover notch 136 may increase
an axial clearance opening 140 and may balance these resistances.
For example, increasing the size of the endcover notch 136 and the
axial clearance opening 140 may decrease the amount of flow
resistance between the axial clearance region 34 and the low
pressure region 38. Further, increasing the size of the endcover
notch 136 and the axial clearance opening 140 may allow a pressure
buildup in the axial clearance region 34 that increases the gap 132
between the rotor and the endcover, thereby increasing the
restoring force of the hydrostatic bearing system 12. As noted
above, increasing the restoring force may provide a stiffer
bearing, which may result in a hydrostatic bearing system 12 with
additional load bearing features.
[0054] It should be noted that in certain embodiments, the endcover
notch 136 may be disposed in any location along the interface of
the rotor 44 and the endcover 31, 33, and can be configured in a
plurality of sizes or shapes. For example, in certain embodiments,
a plurality of endcover notches 136 (e.g., 2, 3, 4, 5, 6, 7, 8, or
more) may be disposed along the interface between the rotor 44 and
the endcover 31, 33. Further, the endcover notches 136 may be
configured in a plurality of shapes (e.g., circle, oval, irregular,
wavy, etc.). Indeed, as noted above, the physical characteristics
of the endcover notches 136 may be determined and optimized based
on the amount of load capacity or stiffness desired from the
hydrostatic bearing system 12.
[0055] FIG. 13 is a schematic diagram of an embodiment of the
hydraulic energy transfer system of FIG. 1, illustrating the
hydrostatic bearing system 12 with the rotor notch 144 disposed
within the axial bearing region 28. Specifically, in the
illustrated embodiment, the axial clearance opening 140 may be
formed by forming a rotor notch 144 along a portion of the rotor
44, such as along the interface between the rotor 44 and the
endcover 31, 33. The rotor notch 136 and the axial clearance
opening 140 may also be configured to spatially increase the axial
bearing region 28 of the hydraulic energy transfer system 10,
thereby resulting in an increased restoring force. The restoring
force acting on the rotor may be configured to increase the axial
clearance region 34 or gap between the rotor 44 and the endcover
31, 33. Further, the greater restoring force acting on the rotor 44
may result in a stiffer hydrostatic bearing system, which may help
provide a greater axial load capacity for the IPX system 20.
[0056] In certain embodiments, similar to the endcover notch 136
described with respect to FIG. 11, the physical characteristics of
the rotor notch 144 may be determined and optimized based on the
amount of load bearing capacity or stiffness desired from the
hydrostatic bearing system 12. Indeed, the rotor notch 144 may be
disposed in any location along the second surface of the rotor 44,
and can be configured in a plurality of sizes or shapes. For
example, in certain embodiments, a plurality of rotor notches
(e.g., 2, 3, 4, 5, 6, 7, 8, or more) may be disposed along the
interface 138 between the rotor 44 and the endcover 31, 33.
Further, the rotor notches 144 may be configured in a plurality of
shapes (e.g., circle, oval, irregular, wavy, etc.).
[0057] FIG. 14 is a perspective view of the end cover 31 having the
axial bearings 118 distributed over the face 122 adjacent to the
rotor 44. The end cover 31 includes an annular groove 156 on the
face 122. In certain embodiments, the annular groove 156 is
configured to receive the HP first fluid 94 to form a fluidic
bearing (e.g., lubricating fluid layer) between the end cover 31
and the rotor 44. The annular groove 156 extends circumferentially
about the axis 158 of the face 234 of the end cover 31.
Additionally, in other embodiments, the axial bearings 118 may
replace the annular groove 156.
[0058] In the illustrated embodiment, a first plurality of axial
bearings 118 are distributed about the axis 158 of the face 122.
The illustrated embodiment includes eight axial bearings 118.
However, in other embodiments, 1, 2, 3, 4, 5, 6, 7, 9, 10, or any
suitable number of axial bearings 118 may be included. As shown,
the axial bearings 118 are radially displaced from the axis 158
along a radial axis 146. In the illustrated embodiment, the axial
bearings 118 are radially offset from the annular groove 156. The
axial bearings 118 include pockets 160 having a circumferential arc
length about the axis 158. In the illustrated embodiment, the axial
bearings 118 are substantially equally spaced from one another
about a circumferential axis 148, forming a segmented bearing.
However, in other embodiments, the axial bearings 118 may not be
equally spaced. For example, the spacing between the axial bearings
118 may be smaller in areas to increase the force on the rotor.
Additionally, the axial bearings 118 may be joined to form an
annular bearing. In the illustrated embodiment, the pockets 160
include a variety of arc lengths (e.g., circumferential extents).
Additionally, in the illustrated embodiment, the pockets 160
proximate the inlet apertures 162, 164 are longer (e.g., have a
larger circumferential extent) than the pockets 160 radially offset
from the apertures 162, 164. However, in other embodiments, the
pockets 160 may all have the same arc length. Furthermore, the arc
length, the radial extent (e.g., width), the depth, or any other
dimension of the pockets 160 may be particularly selected to adjust
for the operating conditions of the rotary IPX 20. In some
embodiments, the face 122 may include a second plurality of axial
bearings 118 radially inset from the first plurality of axial
bearings 118. The second plurality may also vary in arc length,
radial extent, depth, or any other dimension. As will be
appreciated, the location of the axial bearings 118 on the face 122
may be particularly selected based on the operating conditions of
the rotary IPX 20.
[0059] In operation, the axial bearings 118 are configured to use
the pressure of the first fluid 94 to apply a force against the
rotor 44 to eliminate and/or substantially reduce the likelihood
that the rotor 44 contacts the end cover 31, 33. For example, as
the rotor 44 is driven toward the end cover 31, 33 via the first
force 100, the localized pressure in the axial bearings 118
increases, thereby generating a force 124 on the rotor 44 that is
greater than the force 100 on the rotor 44. As a result, the force
produced by the axial bearings 118 maintains the rotor 44 in the
balanced position.
[0060] In the illustrated embodiment, the axial bearings 118 also
include outlets 166 (e.g., axial outlet ports, or openings)
disposed within the pockets 160. While the illustrated embodiment
includes 1 outlet 166 within the pockets 160, in other embodiments
there may be 2, 3, 4, 5, 6, or any suitable number of outlets 166
in the pockets 160. In some embodiments, the outlets 252 are
positioned in approximately the center of the pockets 160. However,
in other embodiments, the outlets 166 may be positioned in any
suitable location in the pockets 160. The outlets 166 are
configured to direct the HP first fluid 94 into the pockets 160
from the feedholes 168 (e.g., radial ports or openings) disposed
upon an outer circumference 170 of the end cover 31, 33. As shown,
the feedholes 168 are positioned upstream of the seal groove 172
(e.g., to enable the HP first fluid 94 to enter the feedholes 168
and supply the HP first fluid 94 to the pockets 168). As will be
described above, the flow passages 128 are configured to direct the
HP first fluid 94 from the feedholes 168 to the outlets 166,
thereby filling the pockets 160 to provide a volume of lubricating
fluid.
[0061] FIG. 15 is a cross-sectional view of the end cover 31 taken
along the line 15-15. As shown, the feedholes 168 extend radially
into the end cover 31, 33 and fluidly couple to the flow passages
128 to direct the HP first fluid 94 toward the outlets 166. In
certain embodiments, each outlet 166 is fluidly coupled to single
feedhole 168 via the flow passage 128. In other words, in some
embodiments, the feedholes 168 are configured to supply the HP
first fluid 94 to only one pocket 160. However, in other
embodiments, the feedholes 168 may be fluidly coupled to multiple
pockets 160 via multiple flow passages 128.
[0062] In certain embodiments, the feedholes 168 may include
inserts or orifices to modify the inlet flow of the HP first fluid
94. For example, the feedholes 168 may include orifices that cause
a pressure drop between the feedholes 168 and the outlets 166. The
pressure drop is configured to enable continuous flow of the HP
first fluid 94. Additionally, the pressure drop may increase the
sensitivity to the pockets 160 to detect small displacements of the
rotor 44. Moreover, in other embodiments, the cross-sectional areas
of the feedholes 168 and/or the flow passages 128 may vary to
modify fluid properties of the HP first fluid 94 as the HP first
fluid 94 is directed toward the axial bearings 118.
[0063] FIG. 16 is a sectional view of the feedhole 168 of the
section 16-16 of FIG. 15. As shown, the feedhole 168 is fluidly
coupled to the flow passage 128 to direct fluid toward the outlet
166 and into the pocket 160. In the illustrated embodiment, the
cross-sectional area of the feedhole 168 is greater than the
cross-sectional area of the flow passage 128. That is, the flow
passage 128 has a converging geometry. In certain embodiments, the
geometry of the flow passage 128 may progressively converge.
However, in other embodiments, the geometry of the flow passage 128
may be continuously converging along the flow passage 128.
Moreover, the feedhole 168 may include a sloped inlet 174
configured to direct the HP first fluid 94 toward a counter-bored
first section 176. Additionally, a second section 178 of the
feedhole 168 may have a smaller cross-sectional area than the first
section 176 of the feedhole 168. Moreover, a third section 180 of
the feedhole 254 may have a smaller cross-sectional area than the
second section 178. As a result, properties of the HP first fluid
94 may be adjusted as the first fluid 94 is directed toward the
outlet 166. For example, as will be appreciated, adjusting the
cross-sectional flow area of the feedhole 168 may cause a drop in
pressure from the feedhole 167 to the outlet 166. As mentioned
above, the pressure drop enables a continuous flow of the HP first
fluid 94 through the feedhole 168. It should be appreciated that
while the illustrated embodiment depicts different cross-sectional
areas of the first, second, and third sections 176, 178, 180, in
other embodiments the cross-sectional areas may be uniform across
the first, second, and third sections 176, 178, 180 of the feedhole
168. Furthermore, the cross-sectional areas of the first, second,
and third sections 176, 178, 180 of the feedhole 168 may also be
adjusted based on operating conditions of the rotary IPX 20.
[0064] The size, shape, etc. of the axial bearings 118 may be
adjusted depending on the expected operating conditions. For
example, the depth of the pockets 160 may be smaller in areas where
the distance 114 is expected to the smaller, thereby increasing the
force 124 produced by the pockets 160 having the smaller depth. In
other embodiments, the end covers 31, 33 a larger number of pockets
160 to increase the force 124 on the rotor 44. Accordingly, the
axial bearings 118 may be designed to substantially reduce and/or
eliminate the likelihood that the rotor 44 contacts the end cover
31, 33.
[0065] FIG. 17 is a perspective view of the sleeve 24 having radial
bearings 120. As mentioned above, due to the first force 100 acting
on the rotor 44, the rotor 44 may tilt, relative to the axis 116.
As a result, the first end 104 of the rotor may be a distance 108
from the sleeve 24, thereby increasing the likelihood that the
rotor 44 may contact or rub against the sleeve 24. The radial
bearings 120 are configured to balance against the force 136
against the rotor 44 to maintain the rotor 44 in the balanced
position. As described above, the HP first fluid 94 enters the
outer channel 130. The radial bearings 120 are positioned along a
length 182 of the sleeve 24 to apply the hydrostatic force along
the length of the rotor 44. In certain embodiments, the radial
bearings 120 are equally spaced along the length 182. However, in
other embodiments, the radial bearings 120 may be positioned on the
sleeve 24 based on anticipated deflection of the rotor 44. For
example, the radial bearings 120 may be positioned near the first
end 104 of the rotor 44 because the first end 104 of the rotor 44
receives direct impact from the first force 100.
[0066] As described above with respect to the axial bearings 120,
the radial bearings include outlets 184 (e.g., radial ports or
openings) fed by the flow passages 128 coupled to the feedholes
168. The outlets 166 direct the first fluid 94 to the pockets 160
to apply the force 136 against the rotor 44. For example, as the
distance 108 between the rotor 44 and the sleeve 24 decreases, the
area between the rotor 44 and the pocket 160 will also decrease. As
a result, the force 136 generated by the radial bearing 120 will
increase, thereby driving the rotor 44 away from the sleeve 24. In
the illustrated embodiment, the radial bearings 120 are disposed on
the interior surface 269 of the sleeve 164. As shown, the radial
bearings 120 may be substantially parallel, substantially
perpendicular, or form an angle with respect to the axis 116.
Moreover, in other embodiments, the radial bearings 120 may be
positioned at any reasonable angle relative to the axis 116.
[0067] FIG. 18 is a partial cross-sectional view of the area 17-17
of FIG. 17 positioned proximate to the housing 56 and the rotor 44.
As shown, the outer channel 130 is configured to supply the first
fluid 94 to the feedhole 168. The first feedhole 168 is disposed on
an outer surface 184 of the sleeve 24. As described above, the
feedhole 168 may include an orifice or insert to restrict or modify
the flow of the first fluid 94. The feedhole 168 is fluidly coupled
to the flow passage 128 to direct the first fluid 94 toward the
pocket 160 via the outlet 166. As shown, the flow passage 128 may
include variances in the cross-sectional area through the wall of
the sleeve 24. For example, the cross-sectional area may decrease
to increase the velocity of the first fluid 94. In other words, the
flow passage 128 may have a converging geometry. That is, the
geometry of the flow passage 128 may progressively converge along
the flow passage 128 or continuously converge along the flow
passage 128. Furthermore, the width, depth, shape, or any other
dimension of the pocket 160 may be modified based on the operating
conditions of the rotary IPX 20. Accordingly, the first fluid 94 is
directed to the pocket 160 to apply the hydrostatic force on the
rotor 44 to maintain the rotor 44 in the balanced position.
[0068] FIG. 19 is a partial cross-sectional view of the area 19-19
of FIG. 17 positioned proximate to the housing 56 and the rotor 44.
In the illustrated embodiment, the feedhole 168 is positioned on an
axial face 186 of the sleeve 24. Additionally, the first fluid 94
is directed toward the feedhole 168, through the flow passage 128,
and into the pocket 160. As mentioned above, the outlet 166 enables
the first fluid 94 to enter the pocket 160 and apply the
hydrostatic force to the rotor 44. Furthermore, as described in
detail above, properties (e.g., cross-sectional area, width, depth,
height, length, etc.) of the feedhole 168, the flow passage 128,
the outlet 166, and the pocket 160 may be adjusted based on the
operating condition of the rotary IPX 20. For example, the flow
passage 128 may include variances in the cross-sectional area. In
other words, the flow passage 128 may have a converging geometry.
That is, the geometry of the flow passage 128 may progressively
converge along the flow passage 128 or continuously converge along
the flow passage 128. Accordingly, the radial bearings 120 may be
configured to apply the hydrostatic pressure against the rotor 44
to maintain the rotor 44 in a balanced position (e.g., the rotor
axis 116 aligned with the axis 74 of the rotary IPX 20).
[0069] FIG. 20 is a plan view diagram of an embodiment of the
endcover 31, 33 of the hydraulic energy transfer system 10 of FIG.
1, illustrating a sink channel 188 and a low pressure sink 190.
Specifically, the first surface of the endcover 31, 33 is depicted
having the high pressure region 36 and the low pressure region 38.
In certain embodiments, a low pressure sink 38 may be provided
about the perimeter of the rotor 44 to utilize the perimeter of the
rotor 44 within the hydrostatic bearing system 12. Specifically,
one or more sink channels 188 may be disposed within the
hydrostatic bearing system 12 to connect the low pressure sink 190
with the low pressure region 38. Indeed, the low pressure sink 190
may be configured to mobilize the perimeter of the rotor 44 to
engage it within the hydrostatic bearing system 12, as well as help
balance the pressure distribution of the IPX system 20. For
example, the low pressure sink loop 196 helps balance the pressures
of the high pressure region 36 and the low pressure region 38
within the IPX system 20, thereby helping to reduce the impact of
the unbalanced pressure distribution on the hydrostatic bearing
system 12 and the IPX system 20, and help to provide additional
bearing capacities and/or axial load capacities for the IPX 20.
[0070] In certain embodiments, the low pressure sink 190 may be
connected to the low pressure region 38 via one or more sink
channels 188. For example, the sink channels 188 may be formed via
a groove 156 disposed in the endcover 31, 33 along the first
surface (e.g., the surface facing the rotor) such that it connects
the low pressure region 38 to the low pressure sink 190. Further,
the sink channels 188 may be configured to act as a drain that
routes the high pressure bearing fluid out of the low pressure
region 38 of the IPX system 20. Accordingly, in certain
embodiments, the high pressure bearing fluid may be configured to
travel from the high pressure region 36 to the low pressure region
38 in the radial direction 146, and may be routed from the low
pressure region 38 to the low pressure sink 190 through one or more
sink channels 188. In this manner, a bearing response is provided
proximal to the low pressure region 38 via the low pressure sink
190, thereby improving the hydrostatic bearing performance within
the low pressure region 38 and through the IPX system 20.
[0071] In the illustrated embodiment, the low pressure sink 190 is
provided as the loop 196 about the perimeter of the rotor 44, with
the sink channel 188 that connects the low pressure sink 190 to the
low pressure region 38 of the IPX system 20. In other embodiments,
one or more sink channels 188 may be configured to connect the low
pressure sink 190 to the low pressure region 38 of the IPX 20, and
the low pressure sink 190 may be a partial loop 198 about the
perimeter of the rotor 44, as further described with respect to
FIGS. 21 and 22.
[0072] FIG. 21 is a cross-sectional diagram plan view of an
embodiment of the endcover 31, 33 of FIG. 20, illustrating a first
sink channel 192, a second sink channel 194, and a partial low
pressure sink loop 198. In certain embodiments, the first sink
channel 192 and the second sink channel 194 may be configured to
connect the low pressure sink 190 with the low pressure region 38
of the IPX 20. Specifically, the first sink channel 192 and the
second sink channel 194 may be configured as a first groove 200 and
a second groove 202 within the first surface 150 of the endcover
31, 33. Particularly, the first groove 200 and the second groove
202 may be configured to form a partial low pressure sink loop 198
around the perimeter of the rotor 44. Indeed, it should be noted
that the perimeter of the low pressure sink 190 may be determined
and optimized based on the magnitude of bearing capacity desired
from the IPX 20. For example, based on the amount of bearing
capacity or axial bearing capacity desired from the hydrostatic
bearing system 12, the perimeter and the number of sink channels
may be determined. Having greater sink capacity will tend to
increase leakage which is undesirable, so there may be a tradeoff
involved.
[0073] FIG. 22 is a plan view diagram of an embodiment of the
endcover 31, 33 of FIG. 20, illustrating the sink channel 188, a
partial low pressure sink 190, and one or more low pressure sink
endpoints 206. In certain embodiments, the low pressure sink 190
may be formed as a loop that starts and ends at the low pressure
region 38 of the IPX 20. In other embodiments, the low pressure
sink 190 may include one or more low pressure sink endpoints 206
configured as a stop point for the low pressure sink loop 196. For
example, in the illustrated embodiment, the low pressure sink 190
begins at the low pressure region 38 of the IPX 20 and includes a
first and a second low pressure sink endpoint 206 that terminates
the loop 196 before it extends across the perimeter of the rotor
44. In this manner, the high pressure bearing fluid that is drained
into the low pressure sink 190 may terminate at the endpoints 206
before reversing the flow direction back towards the low pressure
region 38.
[0074] While the invention may be susceptible to various
modifications and alternative forms, specific embodiments have been
shown by way of example in the drawings and have been described in
detail herein. However, it should be understood that the invention
is not intended to be limited to the particular forms disclosed.
Rather, the invention is to cover all modifications, equivalents,
and alternatives falling within the spirit and scope of the
invention as defined by the following appended claims.
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