EP0761968A1 - Soupape pour moteur à engrenage à denture intérieure avec palier hydrostatique - Google Patents

Soupape pour moteur à engrenage à denture intérieure avec palier hydrostatique Download PDF

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Publication number
EP0761968A1
EP0761968A1 EP95114072A EP95114072A EP0761968A1 EP 0761968 A1 EP0761968 A1 EP 0761968A1 EP 95114072 A EP95114072 A EP 95114072A EP 95114072 A EP95114072 A EP 95114072A EP 0761968 A1 EP0761968 A1 EP 0761968A1
Authority
EP
European Patent Office
Prior art keywords
bearing
rotary piston
control part
pressure
piston machine
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
EP95114072A
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German (de)
English (en)
Inventor
Siegfried A. Dipl.-Ing. Eisenmann
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
Priority to EP95114072A priority Critical patent/EP0761968A1/fr
Priority to US08/702,085 priority patent/US5989001A/en
Priority to JP8234004A priority patent/JPH09177682A/ja
Publication of EP0761968A1 publication Critical patent/EP0761968A1/fr
Ceased legal-status Critical Current

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    • 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
    • F04C2/103—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 one member having simultaneously a rotational movement about its own axis and an orbital movement
    • F04C2/105—Details concerning timing or distribution valves
    • 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
    • F04C2/103—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 one member having simultaneously a rotational movement about its own axis and an orbital movement

Definitions

  • the invention relates to rotary piston machines according to the preamble of claim 1 and to control parts according to claim 13.
  • These rotary piston machines preferably work as slow-running machines with a high pressure torque according to the so-called orbit principle.
  • This primarily means hydrostatic, in particular oil-operated, rotary piston machines.
  • the invention can also be applied to machines which are operated with a compressible working medium, in particular with compressed air.
  • a control part essentially comprises two annular channels that are open to the outside against a contact area. The high pressure and the low pressure side of the two working fluid connections are connected to one channel.
  • connections alternately extend from the two annular channels into a common connection area, from which connection lines lead through a connection part to the displacement part.
  • the annular channels and the connection area are in sliding contact with the lines connected thereto, or with contact surfaces in which the line connections are arranged.
  • the least possible leakage of working fluid under pressure between parts moving against each other or through plain bearings requires the smallest possible gaps between the sliding surfaces. The decency must not be so small, however, that high friction losses and, in particular, high wear and tear occur. It has been shown that the total losses of rotary piston machines largely stem from the losses on the rotating control parts.
  • Drum or butterfly valves are mostly used as control parts.
  • the sliding connection areas and the annular channels are arranged on cylindrical outer surfaces and, in the case of disc valves, at least mostly on the flat side surfaces which are normal to the control part axis of rotation. If necessary, an annular channel is also formed along the cylindrical lateral surface in disk valves.
  • the alternating connections to the common connection area preferably lead through the disk and are therefore not arranged in the plain bearing area.
  • the drum valves have a relatively large flow resistance in the rotating state at high speeds due to the associated increase in the turbulence of the fluid flowing through the channels and connections.
  • the storage is preferably carried out by sliding the cylindrical drum outer surface on a cylindrical inner surface of a housing part. If the so-called port-to-port leakage is to be kept small, the running clearance in the housing must be extremely small, preferably less than 0.5 per mille of the drum diameter. Since the cylinder surface interrupted by channels does not have good sliding bearing properties, contact with the wall cannot be avoided. Due to wear and erosion, the running clearance increases very quickly during operation, so that the channel leakage and also the drainage leakage into the machine interior increase quickly.
  • the flat disc end surfaces must be stored optimally or without leakage and friction.
  • the requirement for the radial bearing or for the cylinder surface depends on whether annular channels and sliding connections are also provided there. However, since the alternating connections to the common connection area are not in the area of the cylinder jacket, this can also be done with a radial external bearing, if applicable better sliding bearing properties are achieved than with common drum valves.
  • Butterfly valves are increasingly being used to achieve better efficiency and a longer service life at high operating pressures.
  • the disc valves have a smaller flow resistance when the connections are designed appropriately compared to the drum valves.
  • a compensation piston is provided, for example, which presses the disk valve in both directions of rotation against the connecting part with the connecting lines to the displacement part without play.
  • relatively high undesirable friction losses occur, which amount to up to 12 percent of the theoretical torque.
  • the losses in forward and reverse running are different.
  • the known control parts are mounted hydrodynamically, which means that the friction is particularly high when the rotary piston machine starts up. In the operating state, a smear layer forms in the bearing. In the event of vibrations, due to variable loads or due to the movement of the rotary piston, the sliding surfaces come into direct contact despite the lubricating layer in the bearing.
  • drum or disk valves It has been shown that an unexpectedly large part of the power loss of the rotary piston machines comes from the rotating control parts, which according to their design are referred to as drum or disk valves.
  • the object of the invention is now to design the rotary piston machine with the rotating control parts in such a way that the leakage and friction losses are significantly reduced.
  • a hydrostatic bearing must be provided for mounting the rotatable control part. Between at least a first sliding surface of the control part and an adjoining second sliding surface of a bearing part, pressurized bearing fluid must be introduced, at least in hydrostatic bearing regions, which can flow into a low pressure region.
  • pressurized bearing fluid must be introduced, at least in hydrostatic bearing regions, which can flow into a low pressure region.
  • at least one carrying pocket in the form of a recess is provided in at least one sliding surface.
  • Each carrier bag is surrounded by a bearing web and fed with bearing fluid under pressure through a supply line.
  • An exit gap is formed around the carrying bag between the bearing web and the sliding surface opposite it. The outlet gap is very small, so that only a small flow of bearing fluid from the carrier bag to a low-pressure area occurs.
  • the supply line, the supply with storage fluid and the outlet gap or the storage gap and the width of the webs are designed in such a way that the pressure required for rigid storage can be built up in the carrying bag.
  • the bearing or outflow gap is in the range from 0.1 to 0.5, preferably from 0.25 to 0.35, parts per thousand of the diameter of the hydrostatic bearing.
  • the outflow gap from the carrier pockets is in the range from 0.2 to 1.2, preferably from 0.4 to 1.0, in particular from 0.6 to 0.8 per mille of the axial thickness of the control part.
  • At least two, but preferably at least three carrier bags, in particular as a carrier bag set, are arranged symmetrically with respect to the axis of rotation of the control part.
  • the carrier bags are preferably circumferential and / or optionally radial and / or if necessary, axially extended and cause an optimal isotropically acting rigidity of the hydrostatic bearing.
  • the hydrostatic bearing has the advantage that even with an extremely small bearing gap, direct contact of the sliding surfaces and the associated friction and wear are avoided due to the high rigidity guaranteed by a sufficiently high bearing fluid pressure and a sufficiently large hydrostatic bearing area.
  • the minimum distance between the sliding surfaces minimizes the leakage losses of the working fluid when entering and leaving the control part.
  • the rigidity and load-bearing capacity of a hydrostatic bearing do not depend on the speed, but only on the supply pressure and the size of the effective surfaces of the carrier bags.
  • the hydrostatic bearing ensures that the control part rotates smoothly right from the start. Due to the high rigidity of the bearings, there is no direct contact with the sliding surfaces even when vibrated.
  • the hydrostatic bearing can be used both for a radial bearing between cylindrical sliding surfaces and for an axial bearing between flat sliding surfaces that are normal to the control part's axis of rotation. This means that the hydrostatic bearing can be used for the external bearing of drum valves as well as for the lateral bearing of disc valves.
  • control parts have two annular channels adjoining the sliding surfaces, one of which is always connected to high pressure, the impression could arise that the pressurized channel has the effect of a hydrostatic bearing.
  • the annular channel does not provide any bearing stiffness desired for hydrostatic bearings.
  • Such a ring duct line around a cylinder jacket surface does not under any circumstances result in a radial deflection of the cylinder jacket surface against this surrounding surface due to the load.
  • the pressurized annular channel cannot function as an axial hydrostatic bearing, even with butterfly valves. Because it is connected to further lines for supplying the displacer part, a reduction in the distance between the sliding surfaces does not lead to resilient pressure increases in the channel. In the case of a hydrostatic bearing with a carrier pocket, the bearing fluid can only escape through the outlet gap, so that a reduction in the outlet gap leads to pressure increases in the pocket and thus restoring forces.
  • carrier bags is particularly advantageous for hydrostatic storage.
  • good bearing or oil film rigidity is achieved radially in all directions by at least three carrier pockets which are distributed substantially uniformly along a circular line.
  • the tilting rigidity of an axial bearing is achieved by two sets of support bags spaced in the axial direction, each with at least three support bags.
  • an essentially isotropic mounting and thus a tilting rigidity are achieved by at least three carrying pockets distributed essentially uniformly along a circular line.
  • the working pressure of the rotary piston machine should preferably be used to supply the hydrostatic bearing.
  • Working fluid is used as the bearing fluid.
  • an increased bearing pressure is automatically set, which means that the rigidity and restoring force to the central position of the bearing of the control part increases with the working pressure and thus with the load on the rotary valve.
  • the oil throughputs through the bearing are extremely low and can hardly be measured.
  • the oil throughput can also be influenced by the dimensioning of the pocket web widths.
  • a pressure potential extending around the mean bearing pressure is preferably made accessible.
  • the supply line with a pre-throttle, the bearing gap and the effective bearing surface of the carrier bag are dimensioned such that the mean bearing pressure is approximately in the range from 1/4 to 3/4, but preferably from 1/3 to 2/3, in particular from substantially 1/2 of the feed pressure.
  • the feeding pressure corresponds to the working pressure or the drive high pressure.
  • the hydrostatic bearing is calculated extremely reliably according to Hagen-Poiseuille's law, assuming a laminar flow. Since both hydraulic resistances, namely that of the pre-throttle and that of the pocket drainage webs or outlet gaps, are linearly viscosity-dependent in the same way, the bearing works at any operating viscosity and thus at any operating temperature.
  • the use of the pressure potential has the advantage that when the bearing is deflected due to the load from its central position, opposing pressure adjustments occur in mutually opposite carrying pockets.
  • the pressure increases in one pocket due to the reduced bearing gap and correspondingly in the opposite pocket due to the enlarged bearing gap.
  • Such pressure differences in the carrier bags lead to a return of the bearing to the central position.
  • two carrier bag sets are preferably provided, one of which is connected in both directions of rotation to the high-pressure and one to the low-pressure spaces of the rotary piston machine.
  • Embodiments according to the invention in which the control part is hydrostatically supported by working fluid, can be implemented with little construction effort.
  • the structural measures can be limited to the control part, so that even rotary piston machines according to the prior art can be converted to machines according to the invention merely by replacing the control part.
  • FIG. 1 b shows a rotary piston machine 1 with an input or output shaft 2, which is supported by two tapered roller bearings 4 in both end regions of the machine and can be rotated about a shaft axis 3.
  • the machine 1 On the exit side of the shaft 2, the machine 1 is sealed leak-free to the outside by a sealing ring 5.
  • Leakage oil lines are preferably provided to relieve the pressure on the seal 5.
  • a leakage oil line 7 is shown, for example, in a first housing part 8 adjoining the end cover 6. If necessary, the leak oil line 7 is also connected to the low pressure side of the working fluid guide via a check valve.
  • a second housing part or connecting part 9 adjoining the first housing part 8 there are connecting lines opening into the displacer part 10 13 provided.
  • a third housing part 11 and a closing part 12 holding the sealing device 5 are arranged between the displacer part 10 and the sealing ring 5.
  • the shaft 2 is provided in the region of the displacer 10 with an external toothing 16 which meshes with the internal toothing 14 of the rotary piston 15.
  • the rotary piston 14 rotates eccentrically around the shaft 2 and meshes with an external toothing 16 in an internal toothing 17 of the displacer housing part 18.
  • FIG. 1 a shows the displacer 10 in cross section and thus gives a good insight into the toothing described.
  • the left half of the working space located between the displacer housing part 18 and the rotary piston 15 must be connected to working fluid under high pressure and the right half must be connected to low pressure at the same time.
  • a control part 19 which is rotatable about the shaft axis 3 and is mounted in the first housing part 8 and in the connecting part 9, is provided in order to ensure the hemispherical feed rotating with the rotary piston 15.
  • the control part 19 is designed as a cylindrical drum valve and comprises on its cylindrical outer surface 20 two annular channels 21 and 22 which are open to the outside. The high-pressure side and the low-pressure side of the two working fluid connections 23 and 24 are connected to one channel 21 or 22.
  • connections 25 extend alternately from the two annular channels 21 and 22 into a common connection area 26, from which the connection lines 13 lead through the connecting part 9 to the displacer part 10.
  • eleven connections 25 are connected to both channels 21 and 22. Due to the alternating contacts with the rotation of the control part 19 between the twenty-two connections 25 connected alternately with high and low pressure and the twelve connecting lines 13, the necessary hemispherical feeding of the displacer part occurs.
  • this includes an external toothing 27 at its end facing the displacer 10, which meshes in the internal toothing 14 of the rotary piston 15.
  • Narrow cylindrical first sliding surfaces 28 are attached to the two end regions of the control part 19 and are required for the radial outer bearing of the control part on second sliding surfaces 29 of the parts 8 and 9 adjoining the control part as bearing parts.
  • three first and three second carrying pockets 30 and 31 are preferably arranged in the sliding surfaces 28 of the control part 19.
  • the carrier bags 30 and 31 are relative Turns by 120 arranged symmetrically around the shaft axis. Exit gaps 32 are formed between the second sliding surface 29 and the carrier pocket edges in the first sliding surface 28.
  • the first three carrier bags 30 are each connected to the closest duct 21 or 22 via a throttle line 35 or a groove with an extremely small cross section.
  • the throttle lines 35 and the outlet gaps 32 are dimensioned in the case of a centrally located control part 19 such that approximately half the high pressure builds up in the carrying pocket 30 when high pressure prevails in the channel 21 or 22 to which the throttle line 35 leads.
  • the throttle lines or throttle channels have a depth that is at least five times, expediently a maximum of ten times, but preferably essentially six times as large as the mean bearing gap width or as the optimal distance between the sliding surfaces.
  • the width of the throttle lines is calculated so that the desired pressure, in particular approximately half the high pressure, is achieved in the pocket.
  • the bearing gap is 5 ⁇ m
  • the depth of the throttle channel is 30 ⁇ m
  • the width is 200 ⁇ m.
  • the carrier bags 31 are connected via leads 34 and pre-chokes 35 connected to connections 25, which connections 25 lead to the more distant channel 21, 22 (FIG. 1 c).
  • the feed lines 34 lead in the control part 19 under the closest channel 21, 22 and include one axial and two radial bores.
  • the axial bore made from the control part end face is closed in the region of the end face with a closing part 36, so that the U-shaped feed line 34 does not leak.
  • the adjoining carrier pockets 30 and 31 are alternately connected to the channel 21 and to the channel 22. Since one of the two channels always carries working fluid under high pressure in both directions of rotation, a carrying bag set with three carrying bags 30 or 31 is always pressurized with working fluid in both control end regions. The working and bearing fluid escaping through the bearing gaps is drained from the machine through leak oil lines 7.
  • Fig. 2 shows an embodiment with a disc-shaped control part 19 '.
  • This disc valve 19 ' comprises an annular channel 21 attached to the cylinder surface and extending to a first side surface 37 and an annular channel 22 adjoining the first side surface. From the channels 21 and 22 there are alternately connecting bores 25 to a second side surface 38 or to the circular connection area 26, where they can come into connection with connection lines 13.
  • the disk valve 19 ' is rotated by a driver sleeve 39 at the speed of the rotary piston.
  • the driving sleeve 39 has an external toothing 27 ′ which meshes with the internal toothing of the rotary piston arranged in the displacement part 10 and an engagement end 40 which engages with the disk valve 19 '.
  • the disk valve 19 ' is hydrostatic between the bearing parts 8 and 9 adjoining the side surfaces 37 and 38 stored.
  • 37 carrying pockets 130 and 131 are formed in the first side surface and 38 carrying pockets 230 and 231 in the second side surface.
  • the carrying pockets 130 and 230 or 131 and 231 are connected to the channels 22 and 21 via connections with grooves 35 which are designed as pre-throttles and are located in the side surfaces 37, 38.
  • FIG. 3 c) there are three carrying pockets 230 and 231 in the second side face 38 along a central circle with respect to a rotation by 120 arranged symmetrically.
  • the pre-throttles 35 connect the carrying pockets 230 or 231 directly to the connecting bores 25, which are connected to the channel 22 or 21.
  • the carrying pockets 130, 131 of the first side surface 37 are arranged in the same way with respect to the axis of rotation 3 as the carrying pockets 230, 231 of the second side surface 38.
  • the feed bores 34 ' can be seen in the carrying pockets 131.
  • the carrier bags 130 are supplied from the channel 22 surrounding them. Part of the leakage current also passes from the channel 22 into the carrier bags 130, which, when the outlet gap of the carrier bags 130 is enlarged, prevents or weakens the pressure reduction required for the reset.
  • the radial distance between the channel 22 and the pockets 130 is preferably as large as possible. If necessary, a separating groove connected to the low pressure is provided between the channel 22 and the pockets 130.
  • the throttles 35 between the channel 22 and the pockets 130 must be dispensed with. The feed must take place from the second side surface 38 in accordance with the feeding of the pockets 131.
  • Fig. 4 shows an embodiment in which the bearings 4 of the shaft 2 are arranged directly on both sides of the displacer 10.
  • a further housing part 11a accommodating the other bearing 4 is provided.
  • the control part 19 ' is rotated via an auxiliary gear 41 by the rotation of the shaft 2.
  • the auxiliary gear 41 since the shaft 2 does not rotate at the speed of the rotary piston 15, the auxiliary gear 41 must generate a transmission which compensates for the transmission during the transmission of rotation from the rotary piston 15 to the shaft 2 and thus drives the control part 19 'at the same speed as the rotary piston 15.
  • the auxiliary gear 41 is preferably designed as a rotary piston gear and constructed essentially the same as the gear of the displacement part 10.
  • an external toothing meshes the shaft 2 in an internal toothing of a transmission piston 15 'and an external toothing of the piston 15' in an internal toothing of the connecting part 9.
  • the transmission piston 15 ' rotates at the same speed as the rotary piston 15 of the displacement part 11
  • the rotation of the transmission piston 15 ' is carried out by a transmission sleeve 42 with an external toothing of the transmission sleeve 42 which engages in the internal toothing of the transmission piston 15'.
  • the disk valve 19 'sits firmly on the transmission sleeve 42 and thus rotates at the same speed as the two pistons 15 and 15'.
  • the embodiment according to FIG. 4 has several advantages. It enables the shaft 2 to be supported directly on both sides of the displacement part 10. Furthermore, the shaft toothing 13 can be made the same or even wider than the rotor toothing 14, so that the tooth strength of the shaft 2 is increased.
  • the displacer part 10 and the control part 19 ' are arranged spatially separated and can optionally be opened or dismantled independently of one another. Due to the optimal bearing arrangement for the shaft 2, the shaft end assigned to the control part 19 'rotates essentially round, so that a needle bearing 43 arranged around the shaft 2 can be used for the radial mounting of the transmission sleeve 42 and thus of the disk valve 19'.
  • the axial bearing of the control part 19 ' is hydrostatic. For this purpose, the control part 19 'according to FIG. 3 is designed. Due to the axial hydrostatic and the radial needle bearing, the control part 19 'rotates with an extremely small loss of friction.
  • the hydrostatic bearing of the control part according to the invention can also be used according to FIG. 5 if a control part 119 is fixedly connected to the input or output shaft which rotates synchronously with the rotary piston, in particular is formed integrally therewith.
  • the hydrostatic bearing includes Carrying pockets 30 and 31 which are arranged at least in the two cylindrical end regions of the control part 119. Due to the connection between shaft 2 and control part 119, the hydrostatic bearing acts as a shaft bearing.
  • a propeller shaft 44 is arranged for the transmission of rotation between the displacer part 10 or the rotary piston 115 and the shaft 2 and is connected in a rotationally fixed manner to the parts connected to it at both ends via toothings.
  • the control part 119 is constructed essentially the same as the control part according to FIG. 1 d), but the toothing 27 is not required because the control part 119 is formed in one piece with the shaft 2.
  • the hydrostatic bearing increases the efficiency of the machine significantly compared to versions without hydrostatic bearing.
  • FIG. 6 shows an embodiment in which the shaft 2 is driven by the displacer part 10 or by the rotary piston 215 at the same speed via a first cardan shaft 44 and the control part 219 via a second cardan shaft 45.
  • the control part is constructed according to FIG. 3 and is thus hydrostatically supported.

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  • Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Magnetic Bearings And Hydrostatic Bearings (AREA)
  • Rotary Pumps (AREA)
  • Actuator (AREA)
EP95114072A 1995-09-08 1995-09-08 Soupape pour moteur à engrenage à denture intérieure avec palier hydrostatique Ceased EP0761968A1 (fr)

Priority Applications (3)

Application Number Priority Date Filing Date Title
EP95114072A EP0761968A1 (fr) 1995-09-08 1995-09-08 Soupape pour moteur à engrenage à denture intérieure avec palier hydrostatique
US08/702,085 US5989001A (en) 1995-09-08 1996-08-24 Planetary rotation machine with hydrostatically mounted control part, and control part for this purpose
JP8234004A JPH09177682A (ja) 1995-09-08 1996-09-04 流体圧支持の制御部を備えた回転ピストン機械並びに該制御部

Applications Claiming Priority (1)

Application Number Priority Date Filing Date Title
EP95114072A EP0761968A1 (fr) 1995-09-08 1995-09-08 Soupape pour moteur à engrenage à denture intérieure avec palier hydrostatique

Publications (1)

Publication Number Publication Date
EP0761968A1 true EP0761968A1 (fr) 1997-03-12

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Application Number Title Priority Date Filing Date
EP95114072A Ceased EP0761968A1 (fr) 1995-09-08 1995-09-08 Soupape pour moteur à engrenage à denture intérieure avec palier hydrostatique

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US (1) US5989001A (fr)
EP (1) EP0761968A1 (fr)
JP (1) JPH09177682A (fr)

Cited By (3)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
EP1074739A1 (fr) * 1999-08-03 2001-02-07 Siegfried A. Dipl.-Ing. Eisenmann Machine hydrostatique à piston rotatif
EP1074740A1 (fr) * 1999-08-03 2001-02-07 Siegfried A. Dipl.-Ing. Eisenmann Machine hydrostatique à piston rotatif
US6524087B1 (en) 2000-08-03 2003-02-25 Siegfried A. Eisenmann Hydrostatic planetary rotation machine having an orbiting rotary valve

Families Citing this family (4)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US6343250B1 (en) 2000-05-19 2002-01-29 Caterpillar Inc. Method and apparatus for smoothing the output of a hydrostatic transmission near zero speed
US6832903B2 (en) * 2002-10-08 2004-12-21 Sauer-Danfoss Aps Functionalties of axially movable spool valve
CH701073B1 (de) * 2004-07-22 2010-11-30 Siegfried A Dipl-Ing Eisenmann Hydrostatischer Kreiskolbenmotor.
EP2628952B1 (fr) * 2011-03-09 2022-04-27 Volvo Car Corporation Système de transmission hydraulique

Citations (3)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE3402710A1 (de) * 1984-01-26 1985-08-08 Siegfried Dipl.-Ing. 7960 Aulendorf Eisenmann Hydraulische kreiskolbenmaschine
US4699577A (en) * 1986-05-06 1987-10-13 Parker Hannifin Corporation Internal gear device with improved rotary valve
EP0367046A1 (fr) * 1988-10-24 1990-05-09 Eisenmann, Siegfried A., Dipl.-Ing. Machine hydrostatique à piston rotatif

Family Cites Families (3)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US4159723A (en) * 1978-02-03 1979-07-03 Danfoss A/S Control device for steering apparatus or the like
DE3015551C2 (de) * 1980-04-23 1986-10-23 Mannesmann Rexroth GmbH, 8770 Lohr Kreiskolbenmaschine
DE3030203C2 (de) * 1980-08-09 1983-03-24 Danfoss A/S, 6430 Nordborg Innenachsige Kreiskolbenmaschine

Patent Citations (3)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE3402710A1 (de) * 1984-01-26 1985-08-08 Siegfried Dipl.-Ing. 7960 Aulendorf Eisenmann Hydraulische kreiskolbenmaschine
US4699577A (en) * 1986-05-06 1987-10-13 Parker Hannifin Corporation Internal gear device with improved rotary valve
EP0367046A1 (fr) * 1988-10-24 1990-05-09 Eisenmann, Siegfried A., Dipl.-Ing. Machine hydrostatique à piston rotatif

Cited By (3)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
EP1074739A1 (fr) * 1999-08-03 2001-02-07 Siegfried A. Dipl.-Ing. Eisenmann Machine hydrostatique à piston rotatif
EP1074740A1 (fr) * 1999-08-03 2001-02-07 Siegfried A. Dipl.-Ing. Eisenmann Machine hydrostatique à piston rotatif
US6524087B1 (en) 2000-08-03 2003-02-25 Siegfried A. Eisenmann Hydrostatic planetary rotation machine having an orbiting rotary valve

Also Published As

Publication number Publication date
JPH09177682A (ja) 1997-07-11
US5989001A (en) 1999-11-23

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