US3779674A - High-pressure gear pump - Google Patents

High-pressure gear pump Download PDF

Info

Publication number
US3779674A
US3779674A US00163750A US3779674DA US3779674A US 3779674 A US3779674 A US 3779674A US 00163750 A US00163750 A US 00163750A US 3779674D A US3779674D A US 3779674DA US 3779674 A US3779674 A US 3779674A
Authority
US
United States
Prior art keywords
internal gear
gear
assembly
housing
axial
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.)
Expired - Lifetime
Application number
US00163750A
Other languages
English (en)
Inventor
O Eckerle
R Jung
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.)
Individual
Original Assignee
Individual
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 Individual filed Critical Individual
Application granted granted Critical
Publication of US3779674A publication Critical patent/US3779674A/en
Anticipated expiration legal-status Critical
Expired - Lifetime legal-status Critical Current

Links

Images

Classifications

    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
    • F04C15/00—Component parts, details or accessories of machines, pumps or pumping installations, not provided for in groups F04C2/00 - F04C14/00
    • F04C15/0003—Sealing arrangements in rotary-piston machines or pumps
    • F04C15/0007—Radial sealings for working fluid
    • F04C15/0019—Radial sealing elements specially adapted for intermeshing-engagement type machines or pumps, e.g. gear machines or pumps
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
    • F04C15/00—Component parts, details or accessories of machines, pumps or pumping installations, not provided for in groups F04C2/00 - F04C14/00
    • F04C15/0003—Sealing arrangements in rotary-piston machines or pumps
    • F04C15/0023—Axial sealings for working fluid
    • F04C15/0026—Elements specially adapted for sealing of the lateral faces of intermeshing-engagement type machines or pumps, e.g. gear machines or pumps
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04C—ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
    • F04C2/00—Rotary-piston machines or pumps
    • F04C2/08—Rotary-piston machines or pumps of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing
    • F04C2/10—Rotary-piston machines or pumps of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing of internal-axis type with the outer member having more teeth or tooth-equivalents, e.g. rollers, than the inner member
    • F04C2/101—Rotary-piston machines or pumps of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing of internal-axis type with the outer member having more teeth or tooth-equivalents, e.g. rollers, than the inner member with a crescent-shaped filler element, located between the inner and outer intermeshing members

Definitions

  • ABSTRACT A high-pressure gear pump with a drive pinion, internal gear ring, and filler wedge, where both the axial thrust and radial thrust are substantially compensated; a deflecting axial thrust plate cooperates with an O- ring to define an axial compensation pressure field; a pair of transverse sealing members, either elastic or radially sliding in grooves, cooperates with the adjacent faces of the periphery of the internal gear assembly and of the housing bore to define a radial compensation pressure field.
  • the invention relates to gear pumps, and in particular to high-pressure gear pumps with axial and radial thrust compensation, where a drive pinion engages a driven internal gear member which is substantially hydrostatically balanced, and where a pivotably mounted filler wedge is arranged between the two gears.
  • the invention suggests a highpressure gear pump where the axial thrust compensation is obtained by providing at least one peripherally restrained relatively thin metal plate which is subjected to an axial compensation pressure field arranged in the housing cover, while the radial compensation pressure field is arranged on the periphery of the internal gear ring or on the periphery of the bearing ring which surrounds the internal gear member and deliniated by two sealing blocks which extend over at least the entire width of the gears.
  • FIG. 2 shows the embodiment of FIG. I in an axial cross-section along the line Il-II of FIG. 1;
  • FIG. 4 shows the embodiment of FIG. 3 in an axial cross-section along the line IVIV of FIG. 3;
  • FIG. 5 shows a modified embodiment similar to that of FIG. 3;
  • FIG. 6 represents a further embodiment derived from the embodiment of FIG. 3, shown in a radial crosssection along the line VII3 VI of FIG. 7;
  • FIG. 7 shows the embodiment of FIG. 6 in an axial cross-section taken along the line VII-VII of FIG. 6;
  • FIG. 8 represents a still further embodiment of the invention shown in a radial cross-section taken along the line VIIIVIII of FIG. 11;
  • FIG. 9 shows an enlarged detail of the gear pump of FIG. 8.
  • FIG. 10 shows the detail of FIG. 9, viewed in the direction of arrow X;
  • FIG. 11 shows the embodiment of FIG. 8 in an axial cross-section taken along the line XIXI of FIG. 8;
  • FIGS. Ill and 12 show, respectively, cross-sectional details of FIG. 11;
  • FIG. 14 represents a still further embodiment of the invention, shown in an axial cross-section.
  • the first embodiment of the invention shows a particularily simple version of the high-pressure gear pump of the invention.
  • the drive shaft I is rotatably supported in the cover plates 2 and 3, and the pinion 4 is in driving connection with the shaft 1.
  • a filler wedge 5 is pivotably mounted by means of a dowel pin 6 which is positioned for rotation in oppositely arranged bores in the cover plates 2 and 3.
  • the internal gear ring 7 is so arranged that a clearance remains between it and the inner diameter of the housing 8.
  • a flexible thrust plate 9 In the separation plane between the central housing 8 and the cover plate 2 is arranged a flexible thrust plate 9.
  • the various pump elements 2, 3, 8, and 9 are firmly clamped against one another by means of bolts 10.
  • the inner space of the gear pump is sealed off against the outside by means of O-rings Ill.
  • Another O-ring l2 deliniates the axial compensation pressure field 13 which receives its pressure fluid through the bore 14.
  • the radial compensation pressure field is deliniated by two sealing blades 16 and I7 which, when no pressure is present, are pushed against the periphery of the internal gear member 7 by means of springs 18.
  • the pressure medium flows into the pump via the intake connection I9 and leaves the pump through the pressure connection 20.
  • the diagonal bores 21a and 21b allow the oil from the bearings to return to the intake side.
  • the internal gear ring 22 is supported inside a bearing ring 23 which is retained by means of two dowel pins 24 and 25 arranged in the plane determined by the two axes of the gears.
  • the bearing ring 23 has its mobility restricted in the direction perpendicular to the above plane in order to maintain a fixed center distance between the gears 4 and 7.
  • two grooves containing two sealing blocks 26 of either rubber or plastic material are arranged along the end lines of the compensation pressure field, either in the periphery of the bearing ring 23 or in the adjacent surface of the housing bore, the unmounted blocks being axially longer than the width of the bearing ring 23.
  • a recess 27 which improves the pressure distribution in the compensation pressure field.
  • the filler wedge 29 can wear during running-in in accordance with the distance e between the dowel pin 24 and the groove by which it retains the bearing ring.
  • This clearance further insures that, under deflection of the pinion shaft 30 and corresponding deflection of the dowel pin 24, the elements 22 and 29, and 23 can follow this displacement without restriction, meaning that the bearing ring 23 must not abut against the inner wall of the central housing 28.
  • the dowel pin must be of such a diameter that its deflection corresponds to the deflection of the pinion shaft.
  • FIG. 5 A modification of the previously described embodiment is shown in FIG. 5, where the dowel pins 31 and 32 are arranged in radial alignment and include a shoulder for the positioning of the bearing ring 33. In this case, it is necessary that the dowel pins are sealed against the outside.
  • This embodiment has the advantage of providing a more stable bearing ring 33.
  • the central housing 34 includes oppositely arranged guide noses 35 which engage matching grooves 36 in the bearing ring 37, so as to restrict the beearing ring motion to the horizontal direction only.
  • a flexible dowel pin 38 At a 90 angle to these guide noses is arranged a flexible dowel pin 38 whose cross section and restrained length is selected to correspond with the elasticity of the pinion shaft 39.
  • the dowel pin, or dowel pins are of such a cross-section and length that their elasticity gives them a deflection under pressure increase which is identical to the deflections of the pinion shaft and of the filler wedge pin.
  • FIGS. 8 through 11 A still further embodiment is shown in FIGS. 8 through 11.
  • the internal gear member is represented schematically only by a chain line, while the filler wedge and the drive pinion are omitted, in order to show the structural details of the thrust plate 41.
  • the oval contour of the bore 42 in the housing 43 can be obtained in a simple manner, for example, by first producing a circular bore which is offcenter with respect to the center of the internal gear member by the amount of a with the radius R, and by producing the same bore oppositely off-center by the amount of a with the same radius R.
  • Another simple manner of producing such a bore is by broaching.
  • the broaching operation in this case, has the further advantage that the guide slots 48 for the sealing blocks 49 can be produced in one and the same operation.
  • a further inventive contribution consists in providing for the thin thrust plate 41 to be die-cut, whereby all the control contours are obtained in a simple blanking operation.
  • pressure-compensated gear pumps it is particularly important that the pressure areas be clearly deliniated, so as to avoid the possibility that, under varying operational conditions, the areas which are subjected to the operational pressure should increase or decrease.
  • the two passages 50 prevent the possibility that the pressure between the tooth gaps increases in this area already, as these passages connect the tooth gaps to the intake side of the pump.
  • Another pair of passages 51 assure that the tooth gaps are clearly communicating at this point with the pressure area in the axial pressure field 52 which is deliniated in FIG. 8 by a broken line.
  • a further pair of small nozzle passages 53 is arranged between the above passages, so as to insure that the pressure increases gradually over a distance somewhat less than one tooth pitch.
  • FIG. 8 also shows a particularly advantageous arrangement of the sealing elements which are represented by the blades 16 and 17 in FIG. I.
  • these blades are replaced by sealing blocks 49.
  • FIGS. 9 and 10 these sealing blocks 49 are illustrated at an enlarged scale, where it can be seen that each pair of sealing blocks is supplied with fluid from the axial compensation pressure field through a die-cut bore 58 in the face plate 41.
  • the pressure fluid passes through the passages 51 and 55, and via the bore 58, into the space delineated by the sealing blocks 49 and by the adjacent faces of the outer periphery of the internal gear member and of the housing bore 42 with the radius R.
  • the sealing blocks 49 include a connecting groove 59 through which the pressure fluid passes into the pressure-balancing recess 60 and, via the bores 61, into the space 62 provided between the sealing block 49 and the housing 43.
  • This arrangement assures that the sealing blocks 49 are always pressed against the internal gear member. In the idling condition, however, when the operational pressure is zero, this hydrostatic force resulting from the overcompensation is not effective.
  • additional leaf springs 63 are provided in the space between the sealing blocks 49 and the housing 43 so that, even during idling, the sealing blocks are pressed against the internal gear ring, while the latter is pressed against the filler wedge, and the filler wedge in turn is pressed against the crown circle of the pinion.
  • each of the sealing blocks 49 is pressed against the lateral retaining face 64, where it forms a clearance-free metallic seal. For this reason, this design does not require any rubber sealing elements.
  • One advantage of this is that the temperature sensitivity of this kind of seal is of no concern.
  • each pair of sealing blocks defines a compensation pressure field of the length L.
  • the space 65 between the two inner sealing blocks is connected to the intake side of the pump via a central groove 66, a peripheral channel 67, and via the end grooves 68.
  • the layout of the compensation pressure field would correspond to that shown in FIG. 1.
  • the latter layout has the disadvantage that the internal gear member is subjected to a very high alternating bending stress. This situation is due to the requirement that the outer compensation pressure field, i.e. the area enclosed between the sealing blades should be only slightly larger than the pressure area on the inside of the internal gear member.
  • FIG. 11 also shows the arrangement of positioning pins 83 which were not shown in FIG. 8.
  • the thrust plate 41 in contrast to the earlier described FIGS. 2, 4 and 7, is not clamped between the housing parts 84 and 43, but has a minute axial clearance inside the recess 69 of the housing part 84. In this recess, the thrust plate 41 is fixed against rotation but free to make a small axial displacement.
  • This arrangement has the advantage that, when temperature differentials are present between either the pinion shaft, the filler wedge, the internal gear assembly, or the housing, such as may occur during start-up at low temperatures, the warmest part of the face plate 41, which is then the thickest, can yield in the direction of the shaft axis, so as to avoid accidental seizure between the thrust plate 41 and the moving parts.
  • FIGS. 11, 12, and 13 is indicated a further inventive contribution which concerns itself with the accurate positioning of the two cover plates 84 and 70 relative to the housing 43.
  • gear pumps where the shaft 71 is supported in two different housing parts, there is always the problem that the two cover plates 84 and 70 must be aligned very accurately relative to one another. Normally, this positioning is obtained by means of dowel pins or by means of tapered pins which require the corresponding bores in the housing parts to be produced at very close tolerances as regards their diameter and their alignment.
  • a well-known procedure to obtain this accurate alignment consists of using a clamping fixture to simultaneously bore all the housing parts, the separate parts being later cleaned of shavings and the pins being mounted during final assembly.
  • the housing or center part 43 is provided on both axial faces with a pair of positioning bores 72 which are slightly tapered and therefore need not be positioned very accurately.
  • the two cover plates 84 and 70 likewise, have each a pair of cylindrical bores which are only roughly predrilled.
  • the parts 43, 84 and 70 are clamped together in a fixture, whereby the latter are aligned with reference to the pinion shaft bore and the bore for the filler wedge pin 74 (see also FIG. 8).
  • a pair of cylindrical pins 83 as shown in FIG.
  • the pins 83 are of a material which can be readily deformed by cold-flowing. Now, when one of the three parts 84, 43, or 70 need replacement, only the deformed pin 83 needs to be forced 'cutand discarded. The other, more expensive housing parts can bereused.
  • FIG. 14 is shown a design which circumvents this problem.
  • This figure shows a pump structure where the central housing 78 has an outer shoulder on each side which engages a corresponding recess in the adjacent cover plates 77 and 78.
  • the bolt 79 only needs to resist the axial load from the compensation pressure field 80.
  • this version requires that thrust plates 81 and 82 are provided on either side of the housing, because the machining of a recessed face in the cover plates with the surface quality necessary for this application requires machining operations which are too costly.
  • a high-pressure gear pump comprising in combination:
  • a closed housing assembly including a housing body with a housing bore, an intake connection for the pumping medium, a discharge connection for same, and at least one removable lateral cover plate;
  • the internal gear assembly arranged inside the housing bore, the internal gear assembly having a diameter which is smaller than the housing bore so as to define a gap therewith which permits limited displacements of the internal gear assembly relative to the housing assembly, the internal gear assembly comprising a rotatable internal gear ring;
  • a drive pinion and drive shaft arranged inside the internal gear ring and mating therewith to create the pumping action, the drive shaft being journalled in the housing assembly;
  • a radially adjustable arcuate filler wedge arranged between the two gears on their pumping side, the narrower end of the wedge delimiting the field within which the gear teeth are exposed to the pumping pressure;
  • At least one radial compensation pressure field arranged in the gap between the adjacent faces of the housing bore and the periphery of the internal gear assembly and communicating with the pumping pressure field so as to overcome the radially outward directed thrust on the internal gear assembly resulting from the pumping pressure field, in order to establish forcible radial contact between the teeth of the internal gear ring and the outer arc of the filler wedge and between the inner arc of the filler wedge and the teeth of the drive pinion,
  • the peripheral length of the radial compensation pressure field, or fields being defined by transverse sealing members which sealin gly interrupt the gap over its entire axial width, the transverse sealing members yielding radially independently of each other in response to changes in the gap opening, when the internal gear assembly changes its position inside the housing bore.
  • the sealing members are in the form of wiper blades which are radially movable in the guide grooves so that they can follow the displacement of the internal gear assembly;
  • the transverse sealing members include means for urging them against the periphery of the internal gear assembly.
  • a gear pump as defined in claim 1 comprising two separate radial compensation pressure fields, the peripheral length of each one being defined by a pair of transverse sealing members, the peripheral gap between the separate radial compensation pressure fields communicating with the intake side of the pump.
  • housing assembly further includes a plurality of centering pins of cold-flowable material engaging adjacent parts of the housing assembly, the latter being thereby permanently centered relative to one another after the pins have been deformed during assembly.
  • the internal gear assembly further includes a bearing ring surrounding and guiding the internal gear ring, the outer periphery of the bearing ring thus being the periphery of the internal gear assembly and forming one of the two adjacent faces which define the gap for the radial compensation pressure field;
  • At least one of these adjacent faces is provided with transverse grooves for the accommodation of the transverse sealing members;
  • transverse sealing members are of a resiliently compressible material
  • the bearing ring includes means for restricting it against rotation while permitting limited displacements of the internal gear assembly relative to the housing assembly and drive pinion.
  • bearing ring restricting means includes in the bearing ring periphery two opposite transverse recesses in alignment with the plane defined by the two gear axes, and two restricting pins secured in the housing assembly and so arranged with respect to the transverse recesses that, while no radial clearance exists between the pins and the bottoms of the recesses so as to maintain a fixed center distance between the gear axes, lateral clearances between the pins and the sides of the recesses permit other limited displacements of the internal gear assembly relative to the housing assembly and drive pinion.
  • the bearing ring restricting means includes in the bearing ring periphery two opposite recesses in alignment with the plane defined by the two gear axes, and two radially aligned restricting pins secured in the housing body, the restricting pins including means for radially abutting against the housing body so adjusted that, while no radial clearance exists between the pin ends and the bottoms of the recesses so as to maintain a fixed center distance between the gear axes, lateral clearances between the pins and the sides of the recesses permit other limited displacements of the internal gear assembly relative to the housing assembly and drive pinion.
  • bearing ring restricting means includes in the bearing ring periphery and housing bore, near their most distant points from the plane defined by the two gear axes, cooperating guide noses and guide grooves whose flanks are perpendicular to the gear axes plane, so as to restrict the displacements of the internal gear assembly to displacements in the direction perpendicular to the gear axes plane.
  • the housing bore is slightly oval in outline, the smallest diameter of the bore being equal to the diameter of the internal gear ring and aligned with the plane defined by the two gear axes.
  • the transverse sealing members are in the form of flat sliding blocks and are radially movable inside the guide grooves so as to follow the displacements of the axis of the internal gear ring.
  • each sliding block has a sliding face for contacting the periphery of the rotating internal gear ring and a back face substantially parallel thereto;
  • each sliding block includes a means for maintaining the contact between its sliding face and the internal gear ring in the absence of hydraulic pressure in the radial compensation pressure field during startup on the pump.
  • housing assembly further includes:
  • the housing assembly further includes a bearing bore inside which the pinion drive shaft is joumalled, and a helical lubricating groove in the bearing bore, one end of the groove communicating with the suction pocket, the other end of the groove communicating with the intake side of the pumps, the direction of the groove helix being such that the rotation of the drive shaft against the groove causes the pumping medium to be frictionally entrained along the groove toward the suction pocket.
  • housing assembly further includes two removable cover plates, one on each side of the housing body;
  • each cover plate has on its inner side an axial recess of a diameter corresponding to the diameter of the thrust plate and of a depth in excess of the thickness of the thrust plate;
  • the housing body has on each side a protruding shoulder which fits into the axial recess of the adjacent cover plate, thereby centering the parts of the housing assembly relative to each other, each shoulder, when positioned in the recess, leaving a slight axial play for the thrust plate;

Landscapes

  • Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Rotary Pumps (AREA)
  • Details And Applications Of Rotary Liquid Pumps (AREA)
US00163750A 1970-07-17 1971-07-19 High-pressure gear pump Expired - Lifetime US3779674A (en)

Applications Claiming Priority (1)

Application Number Priority Date Filing Date Title
DE19702035575 DE2035575A1 (de) 1970-07-17 1970-07-17 Hochdruck-Kleinzahnradpumpe

Publications (1)

Publication Number Publication Date
US3779674A true US3779674A (en) 1973-12-18

Family

ID=5777105

Family Applications (1)

Application Number Title Priority Date Filing Date
US00163750A Expired - Lifetime US3779674A (en) 1970-07-17 1971-07-19 High-pressure gear pump

Country Status (5)

Country Link
US (1) US3779674A (en))
JP (1) JPS5522633B1 (en))
DE (1) DE2035575A1 (en))
FR (1) FR2101659A5 (en))
GB (1) GB1346474A (en))

Cited By (9)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US3912427A (en) * 1973-01-05 1975-10-14 Otto Eckerle High pressure gear pump
US4132514A (en) * 1976-02-16 1979-01-02 Otto Eckerle High pressure hydraulic gear pump or motor
US5197869A (en) * 1991-03-22 1993-03-30 The Gorman-Rupp Company Rotary gear transfer pump having pressure balancing lubrication, bearing and mounting means
EP0987437A1 (de) * 1998-09-15 2000-03-22 Ford Global Technologies, Inc., A subsidiary of Ford Motor Company Innenzahnrad- bzw. Innenkonturrotor-Druckmittelpumpe
US20060193741A1 (en) * 2003-06-03 2006-08-31 Brueninghaus Hydromatik Gmbh Of Eichingen, Germany Gear pump and holding element therefor
US20130071267A1 (en) * 2011-09-17 2013-03-21 Jtekt Corporation Electric oil pump
US20150267702A1 (en) * 2014-02-14 2015-09-24 Starrotor Corporation System and Method for Improved Performance of Gerotor Compressors and Expanders
CN107725357A (zh) * 2017-12-07 2018-02-23 江西应用技术职业学院 一种适用于水液压传动的内啮合齿轮泵
CN109340102A (zh) * 2018-12-03 2019-02-15 湖北海蓝装备科技有限公司 高压内啮合齿轮泵

Families Citing this family (5)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
GB8817284D0 (en) * 1988-07-20 1988-08-24 Jaguar Cars Hydraulic devices
DE19847082B4 (de) * 1997-10-14 2013-01-17 Denso Corporation Drehkolbenpumpe und Bremsvorrichtung, die diese verwendet
ITMI20090017U1 (it) * 2009-01-26 2010-07-27 Fluid O Tech Srl Assieme di separazione tra la camera di alta pressione e la camera di bassa pressione in una pompa volumetrica
CN104074741B (zh) * 2013-03-26 2017-09-29 德昌电机(深圳)有限公司 流体泵
CN110206727A (zh) * 2019-07-02 2019-09-06 潘国陶 一种端面补偿机构及使用该机构的变速器

Citations (5)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US1719640A (en) * 1926-10-30 1929-07-02 James B Tuthill Rotary machine
US1970146A (en) * 1926-03-01 1934-08-14 Myron F Hill Reversible liquid pump
US2132813A (en) * 1933-06-10 1938-10-11 Gunnar A Wahlmark Rotary engine
US3315608A (en) * 1965-08-23 1967-04-25 Eckerle Otto High efficiency wear-compensating gear pump
US3676027A (en) * 1970-03-14 1972-07-11 Hans Molly Crescent machine

Family Cites Families (10)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
CH216223A (de) * 1940-03-30 1941-08-15 Truninger Paul Drehkolbenmaschine.
US3188969A (en) * 1957-09-06 1965-06-15 Robert W Brundage Hydraulic pump or motor
US3034447A (en) * 1959-05-19 1962-05-15 Robert W Brundage Hydraulic pump or motor
DE1266134B (de) * 1960-09-26 1968-04-11 Otto Eckerle Zahnradpumpe
DE1403923A1 (de) * 1960-09-26 1969-11-06 Oilenergetic Establishment Radialdurchstroemte Hochleistungs-Zahnradpumpe oder Motor
DE1553015B2 (de) * 1963-04-11 1977-04-21 Eckerle, Otto, 7502 Maisch Verschleissausgleichende, innenachsige zahnradpumpe
GB1229621A (en)) * 1967-04-18 1971-04-28
GB1233376A (en)) * 1967-11-17 1971-05-26
DE1653871C3 (de) * 1968-01-18 1982-01-21 Sperry Corp., Troy, Mich. Zahnradpumpe oder -motor
DE1801825A1 (de) * 1968-10-08 1970-06-04 Eckerle Otto Hochdruckinnenzahnradpumpe

Patent Citations (5)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US1970146A (en) * 1926-03-01 1934-08-14 Myron F Hill Reversible liquid pump
US1719640A (en) * 1926-10-30 1929-07-02 James B Tuthill Rotary machine
US2132813A (en) * 1933-06-10 1938-10-11 Gunnar A Wahlmark Rotary engine
US3315608A (en) * 1965-08-23 1967-04-25 Eckerle Otto High efficiency wear-compensating gear pump
US3676027A (en) * 1970-03-14 1972-07-11 Hans Molly Crescent machine

Cited By (13)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US3912427A (en) * 1973-01-05 1975-10-14 Otto Eckerle High pressure gear pump
US4132514A (en) * 1976-02-16 1979-01-02 Otto Eckerle High pressure hydraulic gear pump or motor
US5197869A (en) * 1991-03-22 1993-03-30 The Gorman-Rupp Company Rotary gear transfer pump having pressure balancing lubrication, bearing and mounting means
EP0987437A1 (de) * 1998-09-15 2000-03-22 Ford Global Technologies, Inc., A subsidiary of Ford Motor Company Innenzahnrad- bzw. Innenkonturrotor-Druckmittelpumpe
US20060193741A1 (en) * 2003-06-03 2006-08-31 Brueninghaus Hydromatik Gmbh Of Eichingen, Germany Gear pump and holding element therefor
US7413424B2 (en) * 2003-06-03 2008-08-19 Brueninghaus Hydromatik Gmbh Gear pump and holding element therefor
US20130071267A1 (en) * 2011-09-17 2013-03-21 Jtekt Corporation Electric oil pump
US9334862B2 (en) * 2011-09-17 2016-05-10 Jtekt Corporation Electric oil pump with discharge pressure stabilization
US20150267702A1 (en) * 2014-02-14 2015-09-24 Starrotor Corporation System and Method for Improved Performance of Gerotor Compressors and Expanders
US9657734B2 (en) * 2014-02-14 2017-05-23 Starrotor Corporation Gerotor with reduced leakage
CN107725357A (zh) * 2017-12-07 2018-02-23 江西应用技术职业学院 一种适用于水液压传动的内啮合齿轮泵
CN107725357B (zh) * 2017-12-07 2019-06-25 江西应用技术职业学院 一种适用于水液压传动的内啮合齿轮泵
CN109340102A (zh) * 2018-12-03 2019-02-15 湖北海蓝装备科技有限公司 高压内啮合齿轮泵

Also Published As

Publication number Publication date
DE2035575A1 (de) 1972-01-27
GB1346474A (en) 1974-02-13
JPS5522633B1 (en)) 1980-06-18
DE2035575C2 (en)) 1987-08-06
FR2101659A5 (en)) 1972-03-31

Similar Documents

Publication Publication Date Title
US3779674A (en) High-pressure gear pump
US2932254A (en) Gear pump
US3036434A (en) Thrust bearings for hydrostatic transmissions
US2371081A (en) Vane pump
US5009583A (en) Shaft seal and bearing members for a rotary screw compressor
GB1273246A (en) Pressure loaded gear pump or motor
US2816512A (en) Pressure loaded gear pump
US2571377A (en) Rotary displacement pump
US3654761A (en) Fluid handling device with radially variable working chambers
US3371615A (en) Pressure loaded pump
US3073251A (en) Hydraulic machines
US2962972A (en) Power transmission
US3315609A (en) Wear-compensating high efficiency gear pump
US4177025A (en) High-pressure rotary fluid-displacing machine
US2660958A (en) Pressure loaded gear pump
US2864315A (en) Liquid pump
JP2000205142A (ja) 液圧式容積形機械、特に容積形ポンプ
US3043230A (en) High pressure gear pump
US3015282A (en) Pump
US4573889A (en) Bearings for gear pumps
GB1237739A (en) Improvements in or relating to power transmission
US3363578A (en) Gear pump and thrust plate therefor
US3567350A (en) Power transmission
US6716011B2 (en) Hydraulic pump utilizing floating shafts
US3512906A (en) Gear machine