WO2017160985A1 - Compresseur à gaz haute pression à piston axial - Google Patents

Compresseur à gaz haute pression à piston axial Download PDF

Info

Publication number
WO2017160985A1
WO2017160985A1 PCT/US2017/022512 US2017022512W WO2017160985A1 WO 2017160985 A1 WO2017160985 A1 WO 2017160985A1 US 2017022512 W US2017022512 W US 2017022512W WO 2017160985 A1 WO2017160985 A1 WO 2017160985A1
Authority
WO
WIPO (PCT)
Prior art keywords
wedge
oil
compressor
retainer
valve
Prior art date
Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
Ceased
Application number
PCT/US2017/022512
Other languages
English (en)
Inventor
Craig W. Riediger
Current Assignee (The listed assignees may be inaccurate. Google has not performed a legal analysis and makes no representation or warranty as to the accuracy of the list.)
ECO THERMICS Corp
Original Assignee
ECO THERMICS Corp
Priority date (The priority date is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the date listed.)
Filing date
Publication date
Application filed by ECO THERMICS Corp filed Critical ECO THERMICS Corp
Publication of WO2017160985A1 publication Critical patent/WO2017160985A1/fr
Anticipated expiration legal-status Critical
Ceased legal-status Critical Current

Links

Classifications

    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
    • F04B27/00—Multi-cylinder pumps specially adapted for elastic fluids and characterised by number or arrangement of cylinders
    • F04B27/08—Multi-cylinder pumps specially adapted for elastic fluids and characterised by number or arrangement of cylinders having cylinders coaxial with, or parallel or inclined to, main shaft axis
    • F04B27/0873—Component parts, e.g. sealings; Manufacturing or assembly thereof
    • F04B27/0895—Component parts, e.g. sealings; Manufacturing or assembly thereof driving means
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
    • F04B27/00—Multi-cylinder pumps specially adapted for elastic fluids and characterised by number or arrangement of cylinders
    • F04B27/08—Multi-cylinder pumps specially adapted for elastic fluids and characterised by number or arrangement of cylinders having cylinders coaxial with, or parallel or inclined to, main shaft axis
    • F04B27/10—Multi-cylinder pumps specially adapted for elastic fluids and characterised by number or arrangement of cylinders having cylinders coaxial with, or parallel or inclined to, main shaft axis having stationary cylinders
    • F04B27/1036—Component parts, details, e.g. sealings, lubrication
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
    • F04B27/00—Multi-cylinder pumps specially adapted for elastic fluids and characterised by number or arrangement of cylinders
    • F04B27/08—Multi-cylinder pumps specially adapted for elastic fluids and characterised by number or arrangement of cylinders having cylinders coaxial with, or parallel or inclined to, main shaft axis
    • F04B27/10—Multi-cylinder pumps specially adapted for elastic fluids and characterised by number or arrangement of cylinders having cylinders coaxial with, or parallel or inclined to, main shaft axis having stationary cylinders
    • F04B27/1036—Component parts, details, e.g. sealings, lubrication
    • F04B27/1054—Actuating elements
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04B—POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS
    • F04B27/00—Multi-cylinder pumps specially adapted for elastic fluids and characterised by number or arrangement of cylinders
    • F04B27/08—Multi-cylinder pumps specially adapted for elastic fluids and characterised by number or arrangement of cylinders having cylinders coaxial with, or parallel or inclined to, main shaft axis
    • F04B27/10—Multi-cylinder pumps specially adapted for elastic fluids and characterised by number or arrangement of cylinders having cylinders coaxial with, or parallel or inclined to, main shaft axis having stationary cylinders
    • F04B27/1036—Component parts, details, e.g. sealings, lubrication
    • F04B27/1054—Actuating elements
    • F04B27/1063—Actuating-element bearing means or driving-axis bearing means
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16N—LUBRICATING
    • F16N7/00—Arrangements for supplying oil or unspecified lubricant from a stationary reservoir or the equivalent in or on the machine or member to be lubricated
    • F16N7/36—Arrangements for supplying oil or unspecified lubricant from a stationary reservoir or the equivalent in or on the machine or member to be lubricated with feed by pumping action of the member to be lubricated or of a shaft of the machine; Centrifugal lubrication
    • F16N7/366—Arrangements for supplying oil or unspecified lubricant from a stationary reservoir or the equivalent in or on the machine or member to be lubricated with feed by pumping action of the member to be lubricated or of a shaft of the machine; Centrifugal lubrication with feed by pumping action of a vertical shaft of the machine

Definitions

  • the invention also discloses an axial piston machine that provides a new oil lubrication system that distributes oil to the machine through the wedge.
  • Axial piston machines have performed various functions as compressors and pumps and have been driven by electric motors, hydraulic motors, and other mechanical methods in various environments and configurations.
  • the mechanics of the geometry presented by Applicant may be applied to advantage in both pumps and compressors; however preferred embodiments of applicant's invention will hereinafter be primarily descriptive of advantages for compressing working refrigerant fluids in a vapor compression cycle. More specifically the preferred embodiment will be directed to a high-pressure gas compressor for using the natural refrigerant C0 2 as the working gas.
  • C0 2 is known to be a very effective solvent, and oil tends to dilute in the presence of this gas. Oil loss and dilution results in reduced oil viscosity, and maintaining adequate lubrication film under bearing loads is a crucial consideration for dynamic parts.
  • Ramifications are conditions that affect the durability and operational efficiency of such and similar compressors in a vapor compression system. This is particularly important for thermodynamic cycles of high pressure vapor compression of C0 2 gas refrigerant to a transcritical state to be used for heat pump and/or refrigeration.
  • Separation of the working refrigerant gas from the lubricant (oil) and segregated exclusion from the external vapor compression system circuit and associated components is highly advantageous.
  • a primary reason is because oil (liquid) is known to coat the walls of heat exchangers reducing the heat transfer efficiency of the thermodynamic cycle, and/or oil pooling in undesirable points of a gas circuit which may reduce the oil in the compressor to critically low levels.
  • External means of oil separation and components for the management and return of oil to the compressor do exist for the separation of oil from the working gas. This is conventionally accomplished outside of the compressor in the system; however, it is advantageous for many reasons to separate oil and gas inside the compressor in the process of operation. This neither requires nor precludes the use of external oil management components in a system.
  • intake valves might oil-can, deform, or fracture, and/or discharge valves might likewise see deformation and/or potential valve backer failure depending on the strength of the backer structure.
  • Direct piston blow-by gas into oil wetted case areas containing dynamic components results in oil entrainment of the working gas by exposure to oil soaked elements and/or large exposed oil sump region(s).
  • the route of the intake gas from compressor inlet through to the intake valve should be maintained as oil-free as possible and facilitate oil separation as opposed to enhancing oil entrainment of the working gas. This route is largely overlooked in regard to oil separation and temperature control of the working gas, reducing the ultimate PAGE: 4
  • An ideal configuration would segregate the intake gas from the internal oil containing and wetted regions of the compressor prior to the intake/compression cycle, and before passing through an intake valve, all the while facilitating the separation of entrained oil in the process.
  • An ideal compressor would accomplish these tasks in either a horizontal or vertical orientation. Therefore, piston blow-by gas into the compressor's oil bearing regions should be minimized and quickly evacuated limiting undue exposure to the lubricant.
  • return intake gas should not necessarily be directed directly into or through oil rich internal areas of a compressor as a main gas passage to the intake valve, as this significantly enhances oil entrainment of the gas.
  • liquid oil and/or refrigerant
  • Compressor failure or damage may result as liquid oil and/or liquid refrigerant produces a hydraulic manifestation which effectively does not allow normal gas compression because of the liquid medium state of the compound.
  • the resultant pumped liquid which is considered incompressible, imparts slamming stress forces to thin reed valve components which may fracture or otherwise deform. For these reasons, it is vastly preferable for internal means of gas/ oil segregation, and gas/ oil separation, to occur within a compressor machine between inlet gas port and prior to entry through an intake valve into a compression chamber.
  • Oil lubricated compressors (excluding self lubricated or sealed lubricated components) must rely on any combination of four means to provide liquid lubrication to working frictional contacting parts, and assure that oil is supplied and replenished adequately when in operation. These means are: (1) an oil pump, (2) splash lube, or (3) oil mist circulation designed to supply lubrication to moving parts and (4) immersed running operation. Operational design of oil lubricated compressors is specifically dictated by gravity as a first and primary consideration. PAGE: 5 For vertical oriented axially motors, pumps, compressors, and other shaft driven devices, pumping lubricant to high areas from an oil reservoir/sump generally requires an oil pump as a conventional option to oil immersed operation which can be inefficient.
  • the oil sump is below the wedge which in effect "hides” the pistons and slipper shoes and internal piston retention support mechanisms from straightforward splash lube operation.
  • Oil can be distributed around the wedge perimeter but a solution has not been found for splash lubricating the inner surface of the wedge and maintaining a continuous even hydrostatic film while the spinning wedge is flinging oil off of those very surfaces in the opposite direction.
  • Compounding the problem is the fact that the wedge is sloped and therefore oil drains off the surface quickly when stopped. Relying on splash lubrication routed around and over the top of the wedge cannot be counted on to coat the wedge sufficiently and consistently for dry starting and long term durability. The reasonable conventional alternative would be to lubricate from the shaft.
  • Applicant' s invention embodies an improved centrifugal wedge that acts as a pump.
  • the wedge structure allows even distribution of hydrodynamic oil over the wedge surface while providing adequate splash lubrication necessary internally above the wedge lubricating the innermost frictional mechanisms. These are the innermost bottom piston surfaces as well as the combined retainer mechanisms.
  • the wedge simultaneously lifts and splashes the oil on its outer perimeter, splash lubricating the wedge and piston bottom outer surfaces. Gravity induced bearing loads and oil pooling must be considered.
  • the determinations of the internal/external compressor structures and component orientations, horizontal or vertical conventionally yields a final configuration to be used only in its singular design orientation. The packaging of final equipment largely depends on the selected compressor and its physical dimensions.
  • hermetic scroll compressors are known to have a relatively small cylindrical footprint requiring a vertical orientation.
  • This condensed footprint with a taller profile defines a vertical minimum space limit for installed equipment which is determined largely by the length of the motor and compressor vertically stacked and coupled within one hermetic "shell".
  • the scroll compressor will not function properly or fail if operated in a horizontal position.
  • the upshot is that taller compressors with vertical orientations may not qualify for use in the design of limited headroom, low profile packaged equipment designed for tight vertical spaces such as interstitial spaces such as above ceilings.
  • taller compressors such as scroll compressors are conducive for application in equipment designed where a small footprint is desired and floor space is premium and vertical space is adequate.
  • crankshaft reciprocating piston compressors conventionally couple with an electric motor in a horizontal orientation and usually employ oil sumps often with oil pumps or splash lube methods providing lubrication to frictional components. This orientation is better suited for low profile, larger footprint equipment packages.
  • Many examples of differing compressors have been designed for use exclusively in either a horizontal, or conversely a vertical operational orientation. In short, an oil lubricated gas compressor is needed that will PAGE: 7
  • the compressor being the highest uppermost component allows the pistons and valving to be above any oil reservoir, and the compressor drive shaft usually points downward for connection to a prime mover. This is quite acceptable for use with an open drive compressor requiring but few compromises, and further allows an oil sump in the compressor exclusive of and outside of the driving mechanics of the compressor. This of course assumes a compressor rotary shaft seal of adequate design to hold the gas pressures and withstand the temperatures and mechanically generated seal friction caused by shaft rotation and sealing elements. A hermetic or semi- hermetic (fully sealed) application with the motor below the compressor would eliminate the rotary shaft seal requirement because gas is sealed within the entire compressor/motor assembly PAGE: 8
  • a fixed position spherical ball nose segment, post, and spring assembly imparts both necessary force to counteract suction piston forces as well as fixing a centering position of the wobbling piston retaining plate.
  • Centering a piston retainer plate in this way is intended to prevent radial misalignment assuring that wobbling piston slipper skirts do not interfere with respective but oversize retainer bore holes in the retainer plate.
  • Applicant' s invention provides a means of centering the retainer plate using the dynamic geometric position of the piston slipper shoes and slipper skirts on the wedge face plane as a centering mechanism.
  • Applicant' s invention provides a means for allowing vertical or horizontal operation and selectable options of preferred open drive or sealed hermetic drive configurations, all embodied in a single oil lubricated axial machine.
  • the preferred embodiment illustrates an axial wobble- plate multi-cylinder compressor allowing either horizontal or vertical orientation incorporating combined improvements including but not limited to, means of: superior lubrication oil/gas segregation in vertical or horizontal orientation, oil/gas segregation/separation in either orientation, oil distribution to frictional surfaces in either orientation, through shaft and load bearing allowing vertical or horizontal stacking and plural arrangements of compressors/motors, PAGE: 10
  • Fig. 1 is a left side cross sectional view of the inventive compressor with the shaft extending through the head of the compressor.
  • Fig. 1 A is a left side cross sectional view of an alternate embodiment of the inventive compressor in which the shaft does not extend through the head of the compressor.
  • Fig. 2A is a top plan view of the cylinder block
  • Fig. 2B is an enlarged view of the encircled area A of Fig. 2A
  • Fig. 3 A is a top plan view of the retainer plate.
  • Fig. 3B is a cross sectional view taken along line 3B-3B of the retainer plate of Fig. 3 A.
  • Fig. 3C is an end view of the retainer plate of Fig. 3 A.
  • Fig. 1 is a left side cross sectional view of the inventive compressor with the shaft extending through the head of the compressor.
  • Fig. 1 A is a left side cross sectional view of an alternate embodiment of the inventive compressor in which the shaft does not extend through the head of the compressor
  • FIG. 3D is an enlarged sectional view of detail area B encircled in Fig. 3B
  • Fig. 4A is a top plan view of the retainer sleeve.
  • Fig. 4B is front elevation view of the retainer sleeve of Fig. 4A.
  • Fig. 4C is a cross sectional view taken along line A-A of Fig. 4A of the retainer sleeve.
  • Fig. 4D is an enlarged view of the detail area B encircled in Fig. 4C illustrating the retainer sleeve nose.
  • Fig. 5 A is a top plan view of the assembled retainer plate and retainer sleeve.
  • Fig. 5B is a front elevation view of the assembled retainer plate and retainer sleeve of Fig. 5 A.
  • Fig. 5C is a cross sectional view taken along line 5C-5C of Fig. 5 A.
  • Fig. 6 A is an illustrative plan view taken perpendicular to the sloped plane of an axial piston retainer plate illustrating conventional center positioning methods in which retainer plate bore hole clearances allows function with orbiting piston slipper skirts.
  • Fig. 6 AA is an illustrative plan view taken perpendicular to the sloped plane of the axial piston retainer plate showing a novel positioning method which uses contact of orbiting piston slipper skirts within their respective retainer plate bore holes.
  • Fig. 6B is an illustrative plan view taken perpendicular to the sloped plane of the axial piston retainer plate showing the engineered location and sizing of piston slipper skirts and retainer plate bore holes which enable the novel retainer plate positioning method.
  • FIG. 6 A is an illustrative plan view taken perpendicular to the sloped plane of an axial piston retainer plate illustrating conventional center positioning methods in which retainer plate bore hole clearances allows function with orbiting piston slipper skirts.
  • Fig. 6 AA is an
  • FIG. 6C is an illustrative plan view taken perpendicular to the sloped plane of the axial piston retainer plate showing piston slipper skirt and retainer plate bore hole positioning in (3) example wedge slope directions.
  • Fig. 7A is a front elevation view of the wedge.
  • Fig. 7B is side elevation view of the wedge.
  • Fig. 7C is a cross sectional view taken along line 7C-7C of Fig. 7B.
  • Fig. 7D is a cross section view taken along line 7D-7D of Fig. 7C.
  • Fig. 8A is a left side elevation view of the compressor housing.
  • Fig. 8B is a front elevation view of the compressor housing.
  • Fig. 8C is a top plan view of the compressor housing.
  • FIG. 9A is a bottom view of the head with front elevation to the right
  • Fig. 9B is an elevation cross sectional view taken along line 9B-9B of Fig. 9A.
  • Fig. 9C is an enlarged view of the encircled area C in Fig. 9B.
  • Fig. 9D is an enlarged view of the encircled area D in Fig. 9C.
  • Fig. 9E is a front elevation view of the head.
  • Fig. 9F is a top view of the head in Fig 9E.
  • Fig. 9G is a plan cross sectional view taken across line 9G-9G of Fig. 9E.
  • Fig. 1 OA is a top plan view of the port plate.
  • Fig. 10B is an end view of the port plate in Fig. 10A
  • Fig. 11 A is a top plan view of the suction reed valve plate and the suction valves.
  • Fig. 1 IB is an enlarged view of the encircled area A in Fig. 11 A illustrating the suction valve configuration.
  • Fig. 9D is an enlarged view of the encircled area D in Fig. 9C.
  • Fig. 9E is a front elevation view of the head.
  • Fig. 9F is a top view of
  • FIG. 11 C is an end view of suction valve plate in Fig.11 A.
  • Fig. 12A a top plan view of the discharge reed valve plate configuration with discharge valves.
  • Fig. 12B is an end view of discharge valve in Fig. 12A.
  • Fig. 12 C is a plan view of the valve assembly looking at the suction valve stacked upon the port plate stacked upon the discharge valve
  • Fig. 13 is a perspective sectional view of the housing gas and oil separation configuration with function notes.
  • Fig. 14A illustrates a vertical open drive with a dry (no oil) bottom motor operation.
  • Fig. 14B illustrates a vertical hermetic drive with bottom motor oil (wet) or dry operation.
  • Fig. 14C illustrates a vertical open drive with dry (no oil) top mount motor operation.
  • Fig. 14D illustrates a vertical hermetic center mount upsized single motor deployed with PAGE: 14
  • FIG. 14E illustrates vertical hermetic center mounted dual compressors with top and/or bottom mounted motors and/or alternate machine(s), which also allows convenient compressor staging.
  • Fig. 14F illustrates horizontal open drive with dry (no oil) motor operation.
  • Fig. 14G illustrates horizontal hermetic drive with center mounted dual compressors and downsized end-mounted motors.
  • Fig. 14H illustrates horizontal hermetic drive with center mount upsized single motor operation, for multiple compressor operation, which also allows convenient compressor staging.
  • Fig. 15 is an external isometric view of an alternate embodiment of a semi-hermetic compressor.
  • Fig. 16 is a cross sectional view of the alternate embodiment of Fig. 15. Fig.
  • FIG. 17 is an enlarged cross section view of the retainer sleeve engaging the retainer plate to push the piston-shoe assembly down during the suction stroke.
  • Fig. 18 is a further enlarged cross section view through valves during suction stroke.
  • Fig. 19 is a similar cross section view through valves during discharge stroke.
  • Fig. 20 is a horizontal cross section view through the suction port to illustrate liquid separation from suction gas.
  • Fig. 21 is a perspective view of the ring-shaped discharge valve also called the omni-ring discharge valve. PAGE: 15 IV. DETAILED DESCRIPTION OF THE FIRST EMBODIMENT Turning first to Fig.
  • FIG. 1 there is illustrated an inventive axial piston machine which can be a compressor, pump or engine but for simplicity will be referred to herein as a gas compressor 1.
  • Fig. 1 shows interior structure and components of compressor 1 which has a case or housing 2, a head end 19, and a base mount end 20.
  • a cylinder block 3 which contains at least 3 pistons/ring assemblies (piston 4) in respective piston cylinder bores 3a.
  • Cylinder block 3 is a sealed fit to, or may be integral with, housing 2 located in central housing cavity 2a.
  • the preferred embodiment shows a sealed fit of cylinder block 3 to housing 2 and assembled as shown bolted securely with cylinder bolts 9 and cylinder holes 9a in a sealed arrangement of conventional means with head 19a.
  • Each piston 4 incorporates piston ball 4a which is partially encompassed (ball swage fit) by a slipper skirt 5a of piston slipper shoe 5 which are illustrated as one-piece part. Shown as a conventional piston and shoe ball/socket swaged arrangement, it should be noted that in an alternate and reversed configuration the ball might be integral with the slipper shoe and the socket within the piston base.
  • Rotating wedge 12 is affixed to, or alternatively integral as one piece with, shaft 11 which is driven by a rotational force supplied by an electric or hydraulic motor or other mechanical means such as illustrated in Figs. 14A-14H.
  • Wobble plate or wedge 12 presents a smooth, flat low friction angled face surface which piston slipper shoe 5 must follow in a sliding function wobbling about piston ball 4a as wedge 12 rotates.
  • Mechanical rotation upstrokes piston 4 from bottom dead center within its respective bore 3a performing a pumping or compression stroke of a fluid through 180° rotation to top dead center.
  • discharge and suction valving is accomplished by valve assembly 10.
  • Shaft 11 extends into or through head 19a centered radially at head end 19 by bearing 22 while rotating wedge 12 is supported radially and axially by bearings 13 and 14 which may be combined as one bearing assembly.
  • Shaft 11 may be driven or drive from either the base end 20 or the head end 19 as illustrated by using an open drive coupling method 38 and protective shroud 40, separate but attachable to shaft 11 and base end 20 respectively.
  • Alternative integrated PAGE 16
  • Fig. 1A shows an alternate embodiment of a compressor 1, which shows the versatility of the invention for economy of scale cost reduction.
  • Rotating wedge 12 is supported radially and axially exclusively by engineered bearings 13a and 14a which may be combined as one bearing assembly allowing a shortened, single ended shaft 11a which does not extend into or through head end 19.
  • Engineered bearing support means does not require an opposing end of shaft 11a to be incorporated or otherwise supported.
  • Head 19aa (Fig 1 A) illustrates a lifting ring pilot hole 19g in lieu of head 19 (Fig 1) bore hole machining for shaft 11 and seal 15a accommodations. Head bolt holes 19h and head bolts 19i thread into housing 2 at 19k (Fig. 8C). Either machined configuration of head 19a or 19aa may be used with compressor embodiment 1A without compromising performance. This describes the primary difference in embodiments of compressor 1 and compressor la in Fig, 1 and Fig. 1A respectively; the following description will refer to compressor operation as compressor 1/1 a unless otherwise specified.
  • Retainer plate 6 of applicant' s invention Fig.1/1 a counteracts piston 4 inertial and suction downstroke forces by applying equal or higher force to piston slipper shoes 5.
  • Retainer plate 6 assures piston slipper shoe 5 fully and evenly contacts wedge 12 completely through suction downstroke by capturing wobbling piston slipper shoe 5 as slipper skirt 5a protrudes through bore hole 6c of retainer plate 6, thus applying said pull force on piston 4 to slipper shoe 5 which slides on wedge 12 face angle when shaft 11/ 1 la rotates.
  • the retainer plate 6 is illustrated in Figs. 3A-3D.
  • time proven axial piston pump downstroke retention methods are exampled by Hugelman U.S. 7,794,212.
  • a spherical ball nose segment, post, and spring assembly performs the dual function of imparting the necessary force to counteract the piston suction forces previously described, as well as perform fixed center positioning of an axial piston retainer plate.
  • the retainer plate has a central opening 6b through which the shaft 11 passes.
  • a ball nose centers the retainer plate and is fixed in the machine.
  • the novel method of piston retention utilizes the dynamic geometric position of piston slipper skirts 5a which remain perpendicular to the wobbling face plane of wedge 12 in a predictable track. If desired, this configuration allows open space in the center of the axial machine allowing drive shaft 11 extension through head 19 end of compressor 1 such as illustrated in Fig. 1. Figs.
  • FIG. 6A, 6AA, 6B and 6C illustrate with exaggerated wedge slope the geometric tracking of the arrangement whereby piston(s) slipper skirt 5a defines a centering position of retainer plate 6 and respective bore holes 6c so as to achieve dynamic center positioning of retention plate 6 PAGE: 18
  • FIG. 6A illustrates conventional slipper skirts (heavy black lines) which orbit in a circular pattern inside retainer plate bore holes (heavy dashed lines).
  • prior art axial piston machines maintain clearance between slipper skirts and retainer plate bore holes when centered by conventional means.
  • Retainer plate bore holes are necessarily made excessively large to maintain clearance and avoid interference with slipper skirts.
  • slipper skirt 5a clearance with retainer plate bore holes 6c has been reduced to substantially zero (minimal clearance) at the outer orbit perimeter of slipper skirts 5a to locate retainer plate 6 by rotational contact with retainer plate bore holes 6c as slipper skirts 5a orbit therein.
  • This allowance is effective because opposing piston slipper skirts 5a inherently center retainer plate 6 both in the direction and perpendicular to the direction of the wedge slope (arrow in center of each illustration).
  • Fig. 6 B illustrates slipper skirt 5a (heavy black lines) orbiting geometry which defines the slipper skirt diameter and the retainer plate bore hole 6c (heavy dashed lines) pitch circle dimension.
  • slipper skirt 5a By sizing these dimensions for near zero clearance, the outer perimeter of slipper skirt 5a orbit circles are coincident with retainer plate 6 bore holes and slipper skirts 5a are allowed to freely wobble about piston ball 4a and potentially rotate while being in constant contact with retainer plate bore holes 6c. This configuration results in extremely low friction and even wear on components. In response to rotation of sloped wedge 12, slipper shoes 5 and retainer plate 6 wobble in relation to the center axis of compressor 1/1 A.
  • FIG. 6C shows representative slipper skirt 5a positions in (3) different retainer plate 6 slope directions.
  • slipper skirt 5a axial piston groupings (5 illustrated) continuously maintain points of contact with respective retainer plate bore holes 6c in opposing directions that hold center position of the retainer plate 6 as it wobbles in response to wedge rotation.
  • Conventional axial hydraulic pump piston retention methods might be attempted in a compressor for the consideration of allowing a shaft through a head porting area.
  • Such configurations have major drawbacks if applied within a gas compressor.
  • Traditional ball nose spring force centering methods create a line contact sliding frictional interface while wobbling around on a ball nose as a wedge is rotated.
  • retainer sleeve 7 is not restrained as to rotation in cylinder central bore 3d (Fig.2A), but could be by conventional securing means.
  • the contacting angles of sleeve nose 7a with retainer rolling ring 6a are equally matched and there is no significant force to drive rotation of retainer sleeve 7.
  • FIGs 5B and 5C show this assembled interface.
  • An alternate embodiment applies the entire angle to only one of these parts 7/7a or 6/6a, consequently there would be no side loading applied to the flat part.
  • central bore 3 d could be virtually eliminated; however, the sleeve would rotate in cylinder central bore 3d due to the vector forces applied by retainer plate 6. If the overall contact angle is unevenly split between parts 7/7a or 6/6a, combinations of retainer plate 6 side loading forces weighed with rolling contact of retainer sleeve 7 and rotation in its cylindrical bore may be optimized with material and lubrication selection to minimize detrimental wear due to sliding and side-loading friction factors. It is also notable that area in the center of the machine is the driest area within housing core cavity 2a due to centrifugal force spinning oil from inner to outer radial locations. Therefore, a low friction rolling action of these centralized components is a significant advantage especially for dry starts.
  • Retainer plate 6 must be strong enough to withstand deformation due to cantilever reactive forces applied by motion of piston 4, as well as providing a novel method of retainer ring 6 center positioning. This benefit would apply to both axial piston compressors and pumps. Lubrication of dynamic frictional components is particularly important when compressing dry and/or solvent gases such as C02, and gravity is a primary consideration, therefore orientation of compressor 1/1 a must be considered. Initial description will describe a vertical orientation of the preferred embodiment shown. Oil lubricated axial devices used as gas compressors are prone to experience dry running in certain conditions, and consideration should be given to where oil pools and drains.
  • Oil sump 16 is a cavity located inside housing 2 at base end 20 which is substantially concentric with axial shaft 11/11 a.
  • Oil drain ports 16a and 16b are integral with oil sump 16, and are used for exit porting of oil to external cooling means if necessary (not shown), and/or oil sampling, and/or draining compressor 1/1 A of lubricant.
  • Oil sump 16 is the source of lubrication PAGE: 21
  • centrifugal oil pump means are known to be employed within spinning drive shafts in various vertically oriented machines
  • applicant's invention is an improvement for upright vertical axial piston compressors. In a vertical position, adequate lubrication of piston slipper shoe 5 at the frictional interface with wedge 12 sloped angle face is a difficult challenge.
  • Applicants invention embodies an improved centrifugal wedge 12 located within the wedge structure that allows even distribution of hydrodynamic oil film over the wedge 12 surface while providing adequate internal splash lubrication necessary above the wedge 12 for lubricating the innermost frictional mechanisms.
  • the wedge 12 further contains a dry start and slow start oil reservoir in the wedge which distributes stored oil on startup. Additionally, the wedge 12 simultaneously spin lifts and splashes oil around its outer and upper perimeter, splash lubricating piston 4 and slipper shoes 5 bottom outer surfaces.
  • the new method effectively introduces centrifugal pumping configuration achieved through the spinning wedge 12, which now has the dual function of a traditional wedge and as a pump.
  • Wedge cavity 12c is oil filled in fluid communication with oil sump 16. In operation, oil enters wedge cavity 12c near the axial center of the machine at a reduced diameter 1 lb of shaft 11 / 11 a, which may or may not exhibit an impeller surface configuration.
  • Wedge riser(s) hole(s) or cavity 12a may also be configured as a combined monolithic structure functional with shaft 11/1 la and wedge 12, or a rotational locking interface exampled as deepened spline or keyway channels or other conduits (not shown) providing one or more oil paths upward inside the rotating wedge member. Oil is centrifugally distributed from wedge oil distributor cavity 12b over wedge 12 sloped face supplying abundant PAGE: 22
  • oil distributor cavity 12b may be contoured to maximum effect for enhancing oil film continuity over the planar wedge 12 surface, as well as determine the "throw" of excess oil that is spun off and up to the inner mechanisms.
  • Oil channel holes (not shown) in communication with oil distributor cavity 12b may be radially drilled to points exiting the sloped face of wedge 12 in line with or inboard near the center of the circumscribed path of slipper shoes 5 if required for additional slipper shoe 5 lubrication.
  • the bottom of wedge distributor cavity 12b may form one or more reservoir pockets 12d. Figs.
  • FIG. 7d and 7c illustrate one or more holes or cavities configured to collect and store oil upon shutdown to achieve a lubricating dry start advantage by reducing time for oil to reach wedge 12 face when rpm is initialized.
  • Gas/oil separation and gas return within central housing cavity 2a is additionally addressed by applicants' invention.
  • Central housing cavity 2a and all adjacent housing cavities must be maintained at operational suction pressure of the machine and not allowed to build pressure within due to gas blow-by. These communicating cavities must be vented back into suction regions in head 19/19a. It is advantageous to vent this gas immediately to reduce mixing exposure to lubricating oil and particularly advantageous to separate this gas from the oil before venting.
  • Vented gas from vent slot 3 c is further routed around cylinder bolt 9, though valve assembly 10 bolt access holes 9b to thread into head 19/19a in hole 9c.
  • Counterbore 19e (Fig. 9c) registers with assembly 10 bolt access holes 9b and continues housing 2 venting path to head suction vent slot 19f (Fig. 9A).
  • shaft oil slinger 1 lb on shaft 11/l la is a close engineered clearance at entrance to retainer sleeve bore 7b.
  • Shaft rotation of shaft oil slinger 11 b acts as a centrifugal oil separator, slinging oil and/or oil foam radially outward to help separate gas from oil and also lubricate the contact interface retainer sleeve nose 7a of retainer sleeve 7 and retainer plate 6 at retainer rolling ring 6a. This is the driest area of the oil wetted core area of the operational machine.
  • Retainer plate rolling ring 6a may or may not be a separate insert or application of selected material applied or affixed to retainer plate 6.
  • Suction refrigerant returned from a vapor compression system may benefit from conditioning gas phase and/or oil separation.
  • Refrigerant returning from such a system may be in a cold liquid or quasi liquid state termed liquid slugs and by-passed oil from a compressor into a system may return with the suction gas.
  • liquid is considered incompressible and functions to hydraulically impact valve components.
  • housing 2 interior cavities which substantially encircle central housing cavity 2a.
  • Return suction gas from a system enters compressor 1/1 a at housing intake port 18a into housing intake manifold 18 and routed generally circumferentially around central housing cavity 2a, expelling through housing suction ports 2b (Fig. 8C) prior to high pressure compression.
  • This routing as illustrated in Fig.13 increases initial suction gas PAGE: 24
  • wall 2c (Fig.1) of housing central cavity 2a provides an oil washed thermal bridge, or alternatively an insulating surface which may be utilized to enhance or resist thermal transfer to the working fluid and/or lubricant cooling.
  • Wall 2c is adjacent to housing intake manifold 18, either of which may or may not be insulated by coating, insert, or other application (not shown).
  • housing intake manifold 18 further maintains substantial segregation of the incoming return gas from oil mist and soaked internal working components within central housing cavity 2a providing low velocity and wall contact dwell time helping separate oil from gas.
  • Housing intake manifold 18 may further contain an oil separation media (not shown or specified). Separated oil thus pools at the bottom of housing intake manifold 18 draining down into housing return oil cavity 17 through weep hole(s) 44 in the structural web between intake manifold 18 into housing return oil cavity 17 as illustrated in Fig. 13. As seen in Fig.
  • the housing intake manifold provides a circular path around the housing 2 to further allow the separation of oil from gas.
  • At least one oil port 17a is provided and used for oil filling and/or returned oil from external cooling means if required, and/or oil level control means installed into port if required.
  • Housing return oil cavity 17 is in fluid communication or integral as a single cavity space with oil sump 16 forming a cavity to establish an oil level, and further provides a space for dissipation of gas from entrained oil and oil foam.
  • Refrigeration suction gas expels from housing suction ports 2b (Fig. 8C) which are registered by conventional means with head intake ports 19b (Fig. 9A) and access head suction manifold 19c (Fig. 9C and 9G) further communicating with suction valve ports 19d (Fig. 9A).
  • Valve assembly 10 consists of a port plate 23 (Fig. 10A,), a suction reed valve plate 24 (Fig. 11 A), and a discharge reed valve plate 25 (Fig 12 A).
  • Port plate 23 clamps circular suction reed valve plate 24 sandwiched and sealed between port plate 23 and cylinder block 3.
  • Suction valves 24a cover and seal port plate 23 and suction valve ports 23a.
  • Suction valves 24a flex open to cylinder 3a allowing refrigerant gas to fill cylinder 3a upon down stroke of pistons 4, but suction valves 24a are limited in travel by suction valve side stops 24b (Fig.
  • Suction reed valve 24a sidestops 24b position gas pressure loads so as to minimize bending at the valve neck clampline. Suction valve sidestops 24b are positioned alongside the suction valve ports 23a so gas loads are reacted with minimal affect on bending at the neck of suction valve 24a near the clampline. Sidestops are in line across the centroid of the ports to best counteract forces and gas loads around the ports. Reaction forces at the tip of a typical valve create undesirable valve back bending which increases fatigue stresses at the clampline where valve neck bending originates.
  • a conventional design utilizing a valve stop at the tip is inadequate for high pressure operation as exampled by C0 2 vapor compression.
  • Center holes 23c in port plate 23, and 24d in suction reed valve plate 24, and 25c in discharge reed valve plate 25 each allow optional shaft 11 thru protrusion.
  • Means for valving discharge of refrigerant gas from piston cylinder 3a clamps discharge reed valve plate 25 (Fig. 12A) sandwiched and sealed between port plate 23 and head 19/19a.
  • Discharge valves 25a cover and seal port plate discharge valve ports 23b (Figs. 10A and 12C).
  • Discharge valves 25a are pressurized to open and discharge gas from cylinder bore 3a upon full compression up-stroke of pistons 4, but are limited in travel by backer cut 19e (Fig. 9D) in head 19/19a.
  • High pressure discharge gas exits discharge port 23b into discharge manifold 19m (Figs. 9C, 9G) from head discharge port holes 19n (Figs. 9A, 9G.)
  • a conventional alternate not shown to housing intake port 18a might locate intake return gas porting directly into head 19 at any convenient radial position.
  • FIG. 19/19a alternative to housing 2 intake port 18a provides a shunt option to using intake manifold 18 directly piping suction gas directly into head 19/ 19a.
  • This alternative intake porting option provides a combination of thermodynamic temperature, liquid separation, piping and operational orientation options. This may be of advantage in horizontal operation. Horizontal operation may be obtained by orienting ports 18a, 17a, 16a in a vertical top position. The oil reservoir becomes the lower half section of compressor 1/1 a. Oil lubrication now occurs through splash oil distributed by rotating wedge 12 dipping thru the oil bath. Cylinder vent slot 3c and housing suction port 2b remain above the oil level. Applicant's invention enables prolific economy of scale manufacturing and a multitude of application deployments.
  • FIG. 14 A-H A motor, shown generally as motor 46, drives the compressor 1, which may be the compressor illustrated in Fig. 1 or Fig. 1A, as applicable for the various arrangements.
  • the various configurations that the invention allows are meant to illustrate the numerous possibilities in which the invention may be utilized.
  • compressor 100 there is a semi-hermetic high pressure axial piston compressor/pump which will be simply referred to as compressor 100.
  • the compressor has many similarities with the compressor 1 of the first embodiment.
  • the compressor 100 has added features for high speed capability and reduced manufacturing complexity in a semi-hermetic configuration as will be described herein.
  • the compressor housing 108 of the compressor 100 is connected by bolts 1 12 to the sump housing 101 and the head 110.
  • Three external parts with two bolted joints for this embodiment versus five external parts and four bolted joints in comparable reciprocating compressors is a manufacturing advantage for this embodiment.
  • There is an electrical power connection 114 where power is brought into the compressor 100 in order to drive it by means of the electric motor.
  • Fig. 16 is a cross sectional view of the compressor 100.
  • a wedge shaft 120 is vertically mounted in the compressor. The bottom of the wedge shaft 120 is supported by a lower bearing 122 mounted in the sump housing 101.
  • the lower bearing 122 may be a ball bearing, or alternatively, a combination of journal and thrust washer bearings to run at high speed, greater than 3600 RPM.
  • the wedge shaft 120 drives a wedge 121 as previously described in the first embodiment.
  • the wedge shaft 120 is driven by the electric motor having motor stator end windings 128, motor stator laminations 130 and a motor rotor 132 that is shrunk fit onto the wedge shaft 120.
  • valve plate 135 and cylinder block 134 are fixed below the head 1 10.
  • a freely articulating shoe 142 is swaged onto the piston ball 140 of each piston.
  • the shoes with attached pistons are assembled in corresponding bores of retainer plate 146.
  • the flange of each shoe is PAGE: 28
  • the retainer sleeve 148 reacts against significant piston inertia and suction forces on the retainer plate 146. That net force is transmitted through a small area around the contact point 144.
  • the contact point 144 between the retainer plate 146 and the retainer sleeve 148 is equally distant from the centerlines of the retainer plate and the retainer sleeve, geometrically defined by a line bisecting the wedge angle formed by the compressor 100 centerline and a plane through piston balls 140, for rolling contact to avoid mechanical losses and failures from sliding at the contact interface that is inherent in hydraulic pump retention mechanisms. This is illustrated and described in the first embodiment.
  • the flat retainer plate 146 is less complicated to manufacture than the conical profile in the first embodiment or the spherical interface in traditional axial piston hydraulic pumps, but with the contact point 144 positioned for rolling contact in both embodiments.
  • the rolling action of the retainer plate 146 on the retainer sleeve 148 is explained in detail in the previous embodiment.
  • the retainer plate is radially restrained by the piston shoes as explained in the first embodiment.
  • the controlled length retainer sleeve 148 rigidly maintains a small axial clearance in the stack of parts between the wedge 121 and the valve plate 135 without a spring. Consequently, speed is not limited by the spring force. Elimination of the spring with associated locating features reduces manufacturing complexity. Friction loss at the shoe-to- wedge interface is less under normal operating conditions.
  • the alternate embodiment also has a modified oil distribution system. As seen in Fig. PAGE: 29
  • an offset passage 160 extending upward from the bottom of the wedge shaft 120.
  • the spinning shaft forces the oil outward and upward along the outer wall of the offset passage 160.
  • an upper offset passage 165 in fluid communication with the offset passage 160 and at a greater offset from the center of the shaft 120 to force the oil upward into the centered large diameter receiving cylinder 166.
  • Centrifugal force distributes oil evenly around the inside diameter and up the walls of the receiving cylinder 166 for uniform distribution over the top edge 162, over conical surface 168 and onto the wedge surface for lubricating shoes as they slide over the wedge 121.
  • both reed valves and traditional ring valves in the suction function flex to open under pressure created by fluid flow through the valve and are closed by the flexing force in the valve as fluid flow decreases.
  • valve 200 In this open position, the valve 200 remains flat and is supported all around its outer edge to resist downward gas forces with low valve stress, even at high speed.
  • An additional benefit is that the suction plate valve 200 stays wide open until the piston passes through bottom dead center (zero piston velocity) so the cylinder pressure can nearly equalize with the pressure entering the valve unlike PAGE: 30
  • a common omni-ring discharge valve 206 illustrated in Figs. 18 and 19 (a single valve part serves all cylinders).
  • the omni-ring discharge valve 206 is radially positioned by a large bore 208 in the valve plate 135.
  • the discharge pressure holds the omni-ring discharge valve 206 closed flatly against the port seat on the top side of the valve plate 135, Fig. 18.
  • the piston 136 moves upward and cylinder pressure overcomes discharge pressure locally at that cylinder port to flex the omni-ring discharge valve 206 open against the bottom of the head 1 10 above that particular port, Fig. 19.
  • the piston stops at top dead center, flow stops, pressures equalize and flexing forces in the omni-ring discharge valve 206 returns the valve to its normal flat shape against the port in the valve plate 135 and the closed omni-ring discharge valve 206 seals as shown in Fig. 18.
  • the manufacturing advantage is that the head 1 10 limits valve travel without independent backer parts and fasteners used in other devices.
  • liquid separation is modified to provide cooling for the semi- hermetic motor.
  • a high velocity mixture of liquid and gas enters the offset, horizontal suction connection 1 16 and flows toward the opposite wall that is curved around the compressor center.
  • Dense liquid flung against the curved wall of the compressor housing 108 collects along the wall, separates from the gas and drops downward due to gravity. Less dense gas may flow all the way around inside the curved wall of the housing, and some may recycle past the suction connection 1 16. Because the passage 212 gradually narrows along this flow path, most of the gas flow is squeezed between and over the end winding wires, which cools the end windings 128, as the gas is pushed upward to the suction gas passage 109 in compressor housing 108 as seen in Fig. 16. Cast features in the housing provide this separation and cooling functionality without added parts or incremental machining cost.
  • the compressor 100 is dynamically balanced. There is a couple created by axial piston inertia forces offset from the compressor centerline. This couple PAGE: 31
  • the electrical box 105 in Figure 15 is integrated into the cast compressor housing 108 to eliminate a joint between a loose electrical box and the housing typical in traditional reciprocating compressors.
  • gas and oil passages 212 and 1 1 1 are integrated into the compressor housing casting to reduce manufacturing complexity.
  • rolling retainer interface, solid retainer sleeve, non-flexing plate valve for suction and dynamic balance enable the technical breakthrough of high capacity from speed above traditional 4-pole motor speed (1800 rpm at 60 hz.). Elements of this embodiment have been tested at 2-pole motor speed (3600 rpm at 60 hz) and with all the elements of this embodiment, higher speed is achievable.
  • the flat retainer plate, solid retainer sleeve, oil distribution system, non-flexing plate valves for suction, omni-ring discharge valve and features integrated into castings reduce part count and complexity relative to traditional concepts in piston compressors/pumps.

Landscapes

  • Engineering & Computer Science (AREA)
  • General Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • Manufacturing & Machinery (AREA)
  • Compressor (AREA)

Abstract

L'invention porte sur une machine axiale ayant un coin qui entraîne une pluralité de pistons et une plaque de maintien pour maintenir les plots en contact avec le coin. Un manchon de maintien met en prise la plaque de maintien dans une prise de roulement continue quand la plaque de maintien change son orientation angulaire par rapport aux pistons pendant le fonctionnement du compresseur. Le manchon de maintien maintient également la plaque de maintien en position sans avoir besoin de ressorts tels que ceux utilisés dans les précédents compresseurs. L'arbre d'entraînement rotatif a un passage décalé le long de son arbre pour aspirer de l'huile à partir du bac le long de l'arbre d'entraînement jusqu'au coin pour distribuer de l'huile sur la surface angulaire du coin. Une plaque de vanne circulaire commande l'écoulement de gaz dans et hors des cylindres.
PCT/US2017/022512 2016-03-17 2017-03-15 Compresseur à gaz haute pression à piston axial Ceased WO2017160985A1 (fr)

Applications Claiming Priority (2)

Application Number Priority Date Filing Date Title
US201662309531P 2016-03-17 2016-03-17
US62/309,531 2016-03-17

Publications (1)

Publication Number Publication Date
WO2017160985A1 true WO2017160985A1 (fr) 2017-09-21

Family

ID=59850936

Family Applications (1)

Application Number Title Priority Date Filing Date
PCT/US2017/022512 Ceased WO2017160985A1 (fr) 2016-03-17 2017-03-15 Compresseur à gaz haute pression à piston axial

Country Status (1)

Country Link
WO (1) WO2017160985A1 (fr)

Cited By (2)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
CN115163669A (zh) * 2022-07-15 2022-10-11 东方电气集团东方电机有限公司 轴承润滑系统运行方法、旋转设备以及计算机可读存储介质
CN119079901A (zh) * 2024-11-08 2024-12-06 杭州飞宝传动科技有限公司 一种丝杆升降机

Citations (5)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US5286173A (en) * 1991-10-23 1994-02-15 Kabushiki Kaisha Toyoda Jidoshokki Seisakusho Coolant gas guiding mechanism in swash plate type compressor
US5380168A (en) * 1993-01-25 1995-01-10 Kabushiki Kaisha Toyoda Jidoshokki Seisakusho Axial multi-piston compressor having rotary valve for allowing residual part of compressed fluid to escape
US6368073B1 (en) * 1997-05-26 2002-04-09 Zexel Corporation Swash plate compressor
US20130259713A1 (en) * 2012-03-30 2013-10-03 Kabushiki Kaisha Toyota Jidoshokki Swash plate type compressor
US20140328702A1 (en) * 2011-12-07 2014-11-06 Eco Thermics Corporation Axial piston high pressure compressor/pump

Patent Citations (5)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US5286173A (en) * 1991-10-23 1994-02-15 Kabushiki Kaisha Toyoda Jidoshokki Seisakusho Coolant gas guiding mechanism in swash plate type compressor
US5380168A (en) * 1993-01-25 1995-01-10 Kabushiki Kaisha Toyoda Jidoshokki Seisakusho Axial multi-piston compressor having rotary valve for allowing residual part of compressed fluid to escape
US6368073B1 (en) * 1997-05-26 2002-04-09 Zexel Corporation Swash plate compressor
US20140328702A1 (en) * 2011-12-07 2014-11-06 Eco Thermics Corporation Axial piston high pressure compressor/pump
US20130259713A1 (en) * 2012-03-30 2013-10-03 Kabushiki Kaisha Toyota Jidoshokki Swash plate type compressor

Cited By (3)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
CN115163669A (zh) * 2022-07-15 2022-10-11 东方电气集团东方电机有限公司 轴承润滑系统运行方法、旋转设备以及计算机可读存储介质
CN115163669B (zh) * 2022-07-15 2023-07-18 东方电气集团东方电机有限公司 轴承润滑系统运行方法、旋转设备以及计算机可读存储介质
CN119079901A (zh) * 2024-11-08 2024-12-06 杭州飞宝传动科技有限公司 一种丝杆升降机

Similar Documents

Publication Publication Date Title
US20140328702A1 (en) Axial piston high pressure compressor/pump
US5314318A (en) Horizontal multi-cylinder rotary compressor
US7789202B2 (en) Orientation and gravity insensitive in-casing oil management system for fluid displacement devices, and methods related thereto
JP7791909B2 (ja) 低圧チャンバーロータリーコンプレッサー及び空気調和機
CN114909291B (zh) 一种压缩机的回油结构、压缩机和空调器
US6929455B2 (en) Horizontal two stage rotary compressor
EP1965022B1 (fr) Machine a fluide rotative et dispositif a cycle de refrigeration
US6862975B2 (en) Apparatus for lubricating piston type compressor
KR20070010082A (ko) 회전식 유체기계
JP3111670B2 (ja) 斜板式圧縮機における冷媒ガス吸入構造
WO2017160985A1 (fr) Compresseur à gaz haute pression à piston axial
US9140249B2 (en) Swash plate type compressor
JP7398289B2 (ja) レシプロ膨張機及びランキンサイクル装置
US10174746B2 (en) Piston drive assembly
CN210050033U (zh) 旋转式压缩机和换热设备
CN217652911U (zh) 一种压缩机的回油结构、压缩机和空调器
US9163620B2 (en) Oil management system for a compressor
US5339779A (en) Piston connection for a reciprocating piston engine
US11835046B2 (en) Main bearing housing assembly and scroll compressor having the main bearing housing assembly
JP7411433B2 (ja) レシプロ膨張機及びランキンサイクル装置
KR101587168B1 (ko) 로터리 압축기
CN110761976B (zh) 压缩机
WO2023139829A1 (fr) Compresseur rotatif
US20220025875A1 (en) Compressor
JPS63280879A (ja) 密閉型電動圧縮機

Legal Events

Date Code Title Description
NENP Non-entry into the national phase

Ref country code: DE

121 Ep: the epo has been informed by wipo that ep was designated in this application

Ref document number: 17767432

Country of ref document: EP

Kind code of ref document: A1

122 Ep: pct application non-entry in european phase

Ref document number: 17767432

Country of ref document: EP

Kind code of ref document: A1