EP1776525A1 - Moteur a piston rotatif hydrostatique - Google Patents

Moteur a piston rotatif hydrostatique

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Publication number
EP1776525A1
EP1776525A1 EP05761659A EP05761659A EP1776525A1 EP 1776525 A1 EP1776525 A1 EP 1776525A1 EP 05761659 A EP05761659 A EP 05761659A EP 05761659 A EP05761659 A EP 05761659A EP 1776525 A1 EP1776525 A1 EP 1776525A1
Authority
EP
European Patent Office
Prior art keywords
shaft
hydrostatic
teeth
rotary
engine according
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.)
Granted
Application number
EP05761659A
Other languages
German (de)
English (en)
Other versions
EP1776525B1 (fr
Inventor
Siegfried A. 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.)
EISENMANN, SIEGFRIED A.
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
Publication of EP1776525A1 publication Critical patent/EP1776525A1/fr
Application granted granted Critical
Publication of EP1776525B1 publication Critical patent/EP1776525B1/fr
Active legal-status Critical Current
Anticipated expiration legal-status Critical

Links

Classifications

    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F03—MACHINES OR ENGINES FOR LIQUIDS; WIND, SPRING, OR WEIGHT MOTORS; PRODUCING MECHANICAL POWER OR A REACTIVE PROPULSIVE THRUST, NOT OTHERWISE PROVIDED FOR
    • F03C—POSITIVE-DISPLACEMENT ENGINES DRIVEN BY LIQUIDS
    • F03C2/00—Rotary-piston engines
    • F03C2/22—Rotary-piston engines of internal-axis type with equidirectional movement of co-operating members at the points of engagement, or with one of the co-operating members being stationary, the inner member having more teeth or tooth- equivalents than the outer member
    • 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

Definitions

  • the invention relates to a hydrostatic, low-speed rotary piston engine according to the preamble of the independent claims 1 and 2.
  • a hydrostatic rotary piston engine of this type is known from EP 1 074 740 Bl.
  • An advantage of the disclosed there formation of a rotary piston engine is compared to previous solutions is that the bearings of the hydrostatically highly loaded part of the shaft are arranged immediately adjacent with small axial distance in the fixed housing, so a small amount of bending and tooth deformation on the shaft and, accordingly, a maximum Printing performance and thus to torque delivery can be achieved. Because of this bearing arrangement, there is no way to create a 1: 1 rotary connection between the acting as a rotor rotary piston and responsible for the commutation rotary valve, it has been proposed to synchronously drive the rotary valve via a gear transmission from the shaft.
  • this gear transmission is an eccentric internal gear, in which the disk-shaped rotary valve itself acts as an eccentric member of this transmission and thus performs an unavoidable orbital movement.
  • the invention has as its object to eliminate these deficiencies and at the same time reduce the caused by the orbital movement slightly increased friction at the rotary valve and the production costs.
  • the invention eliminates these disadvantages while retaining the above-mentioned advantages of such machines.
  • the hydrostatic, low-speed rotary piston engine comprises a power unit acting as an output with a centric fixed stator, a rotary piston as the rotor and a centrically mounted shaft.
  • the stator has an internal toothing with the number of teeth d.
  • the rotary piston has a part in the
  • Internal teeth of the stator engaging external teeth with a number of teeth c and an internal teeth with a number of teeth b.
  • the shaft meshes with its external teeth with a number of teeth a partially the internal toothing of the rotary piston, wherein the rotary piston for performing an orbital movement is arranged and dimensioned eccentrically such that with working fluid ver and disposable tooth chambers between the inner teeth of the stator and the outer teeth of the rotary piston form.
  • An inlet and outlet part is used for supply and disposal of the power unit with the working fluid.
  • the rotary piston engine comprises a toothed gear, which is arranged between a - formed in particular by a sun gear - shaft external teeth of the shaft with a number of teeth w and an internal toothing of a fixed internal gear with a number of teeth z as a synchronous drive for the rotary valve.
  • the shaft is mounted on both sides of the power section immediately adjacent bearings arranged.
  • the gear transmission is arranged exclusively in the leakage oil region of the engine and is arranged by a planetary gear with at least one planet carrier, which is rotatably connected to the rotary valve and on which planet gears between the shaft outer teeth and the stationary inner ring gear are arranged, or preferably by an eccentric with an eccentric , which is rotatably connected to the rotary valve formed.
  • the inventive wobble gear requires a much smaller eccentricity, which according to the invention is independent of the eccentricity of the rotary piston in the power section, so that this wobble angle is substantially smaller than half of that wobble angle of the earlier construction.
  • this wobble angle is substantially smaller than half of that wobble angle of the earlier construction.
  • the particular disc-shaped eccentric When using an eccentric, the particular disc-shaped eccentric is rotatably connected via a cup-shaped connecting part with the rotary valve via Mit supportiveverschwept in the speed ratio of 1: 1.
  • the eccentric has, for example, an internal toothing with a number of teeth x and an external toothing with a number of teeth y and is arranged between the shaft external toothing and the internal toothing of the stationary internal toothed ring, so that the corresponding internal and external teeth mesh with each other in a known manner.
  • the equation expression is a positive integer, preferably equal to 3. It must also be striven that in this area, the diameter of the shaft is sufficiently large, so that their torsional strength for a possibly connected holding brake for the maximum
  • Torque is still sufficient.
  • the eccentricity of the transmission is relatively large, so that the wobble angle is correspondingly large.
  • the speed of eccentricity would be quite small.
  • Ne _ wy Nw xzwy ' where this ratio is preferably between -3 and -9.
  • Ne / Nw then the rotational speed of the eccentricity becomes higher, but still remains below the value of the wobble wave of earlier known constructions.
  • Teeth very large, because the propeller shaft is known to absorb any radial forces. Especially at low speed and high working pressure therefore the friction losses and the wear of the teeth are extremely large. Therefore, the starting efficiency of these machines is correspondingly poor and is only about 63 to 71%.
  • the radial load of the teeth between the rotary piston and the stator is only one
  • the output-side roller bearing requires a higher radial load capacity for additional reception of the wheel load. It should be located as close to the center of the wheel. Since, for example, in material handling equipment shock-like elevations of the static wheel load can occur, it is advantageous if this
  • Bearing is as close as possible to the wheel flange and optionally located outside the leakage space of the rotary engine with a rolling bearing grease permanent filling directly in the housing part of the rotary piston engine.
  • the rotary piston engine according to the invention is due to the advantageous bearing arrangement and the powerful continuous shaft, inter alia, excellent as a wheel motor or winch drive for direct driving a wheel or a cable drum.
  • the shaft is preferably formed integrally with a wheel flange on which a wheel or a cable trench for direct drive can be mounted directly.
  • FIG. 1 shows a first embodiment of a rotary piston engine with an eccentric gear in a longitudinal section along the section line C-C of Fig. 2,
  • 1.2 is a cross section through the eccentric gear of the first embodiment of the rotary piston engine
  • Fig. 2 shows a cross section along the section line D-D of Fig. 1 by the rotor-stator system of the first
  • FIG. 3 shows a cross section through the rotor-stator system of an embodiment with rotationally mounted rollers as internal toothing in the stator
  • FIG. 4 is a view X on Fig. 1 on an SAE connection of an embodiment, a partial section along the line A and a partial section along the line B of Fig. 3,
  • 5 shows a longitudinal section through an embodiment of a wheel motor according to the invention
  • 6 shows a longitudinal section through a wheel motor according to the
  • FIG. 7 shows a longitudinal section through a wheel motor according to the invention with the second coupled to the shaft
  • FIG. 9 shows a possible hydraulic circuit diagram for controlling the 2/3-stage motor according to FIG. 7 and FIG. 8 with exemplary technical data, FIG.
  • 10 is a longitudinal section through an inventive rotary piston engine with a coupled to the shaft, large-dimensioned working brake as a multi-disc brake,
  • 11 is a longitudinal section through an advantageous development of an inventive rotary piston engine with a circumferential axial relief groove on the axial sliding surface between rotary valve and balance piston,
  • Fig. 13 is a longitudinal section through the rotary valve and the
  • FIG. 14 is a left side view of the rotary valve and the balance piston of FIG. 13th
  • possible embodiments with reference to several figures, which partially show a single embodiment in different views with sporadically different levels of detail, explained, reference being made in part to already mentioned in previous figures reference numerals.
  • Fig. 1 shows a first embodiment of an inventive rotary piston engine with a
  • Fig. 2 shows a cross section through the rotor-stator system of the first embodiment along the section line D-D of Fig. 1. Furthermore, in FIG. 2 the cutting direction of FIG. 1 can be seen from the section line C-C in FIG.
  • the rotor-stator system of the power unit 1 of the rotary piston engine comprises a centric fixed stator 4 with an internal toothing 5, hereinafter referred to as first internal toothing 5, into which an eccentrically arranged for performing an orbit movement, acting as a rotor rotary piston 6 with a hereinafter as first external toothing 7 called external teeth at least partially engages.
  • a centrally mounted by means of two on both sides of the power unit 1 immediately adjacent bearings 10, 11 mounted shaft 2 has an external toothing 9 - the second external teeth 9 -, which in turn at least partially engages in an internal toothing 8 of the rotary piston 6, called the second internal toothing 8.
  • the forward direction of rotation of the rotor-stator system of the rotary engine is defined for the following explanations as the direction of rotation in which the
  • the expanding swallow cells lie between the first internal toothing 5 and the first external toothing 7 Always on the left and the compressing feed cells always to the right of an eccentric axis 62. Since the eccentric axis 62 performs one of the direction of rotation 61 of the shaft 2 and the direction of rotation 60 of the rotary piston 6
  • Mirrorsetze direction of rotation 64 creates a rotating field for the radial hydraulic force on the rotary piston 6, if always High pressure is supplied to the expanding swallow cells.
  • the control of this rotating field worried a rotary valve 3 as a commutator, similar to a DC motor.
  • the rotary valve 3 has eleven evenly distributed on the circumference, with the first annulus 56 in communication high-pressure window 21a.
  • a control plate 22 with control slots 21 has twelve uniformly distributed on the circumference pressure window 33 a, which are connected via feed bores 33 with the twelve toothed chambers between the first internal toothing 5 of the stator 4. Because of the circumferential distribution eleven to twelve of the high-pressure window 21a of the rotary valve 3 and the pressure window 33a of the control plate 22 is always only one half of the tooth chambers of the stator 4 under high pressure, and that with the correct phase position of the rotary valve 3 with the rotary piston 6 always those tooth chambers, in of Fig. 2 are left of the eccentric axis 62.
  • the shaft 2 is mounted roller-mounted directly in the housing on the left and right of the rotor-stator system, so that the drive of the rotary valve 3 must take place via the shaft 2, which due to the system performs a different rotational speed than the rotary piston 6.
  • the shaft 2 runs three times as fast about its axis as the rotary piston 6 about its own axis. Accordingly, the rotary engine according to the
  • a transmission between the shaft 2 and the rotary valve 3 with the same translation to slow can be done by means of an eccentric gear 30, as in the first embodiment according to FIG. 1 and FIG. 1.2, or by means of a planetary gear 80, as shown in a second embodiment according to FIG. 1.1.
  • Fig. 1.1 shows the second embodiment of an inventive rotary piston engine with a planetary gear 80 in a partial longitudinal section along the
  • the planetary gear 80 includes a sun gear 13 on the shaft 2, the shaft outer teeth 14 meshes with planetary gears 90, which • are mounted on a planet carrier 91, the 1: 1 with the rotary valve 3 torsionally rigid is coupled.
  • the planet gears 90 mesh simultaneously with a fixed internal gear ring 92, which has twice the number of teeth as the sun gear 13 on the shaft 2. According to the laws of the planetary gear is then the translation of the shaft 2 to the rotary valve 3 exactly 3: 1 slow.
  • a simpler constructed eccentric gear 30 which includes a sun gear 13 on the shaft 2 with a Wellenaussenvertechnikung 14 and a fixed inner ring gear 28, the inner teeth 17, hereinafter fourth internal toothing 17 called, compared to the number of teeth of the shaft outer teeth 14 double the number of teeth.
  • a disk-shaped eccentric 26 which has an internal toothing 15 in the interior - the third internal toothing 15 - and outside an external toothing 16, referred to as the third external toothing 16, has.
  • this eccentric gear 30 is executed with tooth shapes that allow the number of teeth difference between the shaft outer toothing 14 and the third inner toothing 15 and the third outer toothing 16 and the fourth inner toothing 17 is equal to 1.
  • involute teeth such transmissions are usually not feasible, as in this case, dental head interference takes place. Also, they do not allow exact radial centering of the wheels against each other under these conditions. It should therefore be resorted to other tooth shapes.
  • a double-cycloidal internal-external toothing is preferably used, as is known, for example, from German Patent DE 39 38 346, to which reference is hereby made.
  • This eccentric gear 30 also has a reduction ratio between the shaft 2 and the disk-shaped eccentric 26 of exactly 3: 1 in the slow.
  • the disc-shaped eccentric 26 is 1: 1 rotationally connected via a cup-shaped connecting part 27 rotatably connected to the rotary valve 3, wherein Mitauervertechnik Weg 31 and 32 allow the cup-shaped connecting part 27 together with the disc-shaped eccentric 26 a small wobbling movement corresponding to Exzentermony the disc-shaped eccentric 26 performs.
  • the backlash of the shaft outer teeth 14, the third internal teeth 15 of the eccentric 26, the third external teeth 16 of the eccentric 26, the fourth internal teeth 17 of the internal ring gear 28 and the Mit psychologyveriereept 31 and 32 are designed to be slightly larger than usual because of the wobbling motion.
  • an axial compensating piston 65 is provided in a known manner.
  • FIG. 3 shows a cross-section through the rotor-stator system of a further exemplary embodiment, in which rollers 81 mounted in rotation are used as the first internal toothing 5 in the stator 4.
  • These rollers 81 should always be trapped in their cavities 82 in the stator 4, i. the caverns 82 should taper toward the shaft 2 beyond the roller radius so that the rollers 81 can not move radially inward out of the caverns 82. This would lead to a blockage of the rotary engine.
  • the shape of caverns 82 is clearly illustrated.
  • Fig. 3 and in Fig. 4 which shows a view X on a SAE connection, a partial section along the line A and a partial section along the line B of Fig. 3, it is also shown that two of the twelve screws are designed as fitting screws 93, which are to be used in the assembly of the engine as the first. From Fig. 4 is also in the partial section A of Fig. 3 can be seen that the rotary engine should be designed very compact due to the specified by the international SAE standard hole patterns for mounting the engine so that dimensions and weight are optimized.
  • a flange screw connection for the high and low pressure connection 55 or 57 according to SAE standard is also shown here ' .
  • One application for the rotary piston engine according to the invention is the use as a wheel motor, as shown in its simplest form as a longitudinal section in FIG. 5.
  • Extremely advantageous in this embodiment of a wheel motor is the formation of a driven-side roller bearing 11 outside a leakage space 85 directly in the housing part 84 of the engine. Since such wheel motors do not require high speeds, a permanent rolling bearing fat filling is sufficient as lubrication, which is sealed by a NILOS ring 72 to the outside.
  • a wheel flange 40 can be made integral with the shaft 2, so that for large wheel loads, the shaft 2 is very robust auslagbar.
  • a hydrostatic wheel bearing usually requires a spring-loaded, automatically spring-loaded parking brake, which is independent of the hydraulic pressure, in order to prevent the parked vehicle from rolling away.
  • Fig. 6 shows a possible realization of such a wheel motor in longitudinal section, in which on the side opposite the output a spring-loaded parking brake 42 is arranged in the form of a multi-disc brake.
  • the inventive rotary piston engine allows advantageously a form suitable for large torques through shaft 2 with a large 'sized wool extension 41 so that the slats of the parking brake 42 directly via a hub 73 can transmit their braking torque to the shaft 2.
  • the shaft outer toothing 14 for the eccentric gear 30 is extended outwardly in manufacturing technology, on which the hub 73 can be wedged torque-effective torsionally effective.
  • This spring-loaded parking brake 42 is a wet-running multi-disc brake, which can be released with greatly reduced hydraulic pressure via the separate port 43.
  • a plate spring 74 is provided here.
  • Wave extension 41 receives a larger torque capacity. This is of particular importance, in particular with wide moving sets in the power unit 1, as explained below. Since with the broadening of the moving set of the power unit 1 and the torque transmitting second internal teeth 8 of the rotary piston 6 and the second external teeth 9 of the shaft 2 widened automatically, here the high pressure level can be largely maintained and thus an increase in performance can be achieved. In the machines with propeller shaft drive between the rotor and the output shaft, this is not possible. Therefore, there is usually only allowed a lower pressure level at wider Lauf algorithmsn with the stator 4 and the rotary piston 6. Motors with wide Lauf algorithmsn run because of the higher absorption rate usually slower, so that the life of the bearings 10 and 11 is not a major problem.
  • a hydraulic motor is shown in longitudinal section or cross section according to the invention, in which except the first power part 1 on an extended shaft end 44 of the shaft 2 a torsionally rigid coupled to the first power unit 1 second, preferably narrower power unit 46 is arranged with its own radial bearing 47, which can be operated via the ports 75 and 76 separately with working fluid, preferably from one and the same hydraulic pump.
  • a proposal on the control of such a 2/3 stage motor with the first power section 1 and the second power section 46 is shown in Fig. 9 in the form of a hydraulic circuit diagram with exemplary performance data.
  • a throttle valve serves as a brake valve 87, in particular when driving downhill of the vehicle.
  • a valve 86 By means of a valve 86, the operating state of the drive from operation D to neutral N can be switched.
  • a further rotary engine according to the invention is shown in longitudinal section, which can of course also be designed as a wheel motor according to FIG.
  • a hydraulically releasable spring-loaded working brake 50 designed as a disk brake, is arranged on a shaft extension 52.
  • This work brake 50 the braking force is applied by means of springs 78, for example, in a hydrostatically driven winch for car or ship cranes the task of keeping the full permissible Seijlast that corresponds to the maximum pressure and thus the highest torque of the engine in the balance , without support hydraulic pressure on the engine.
  • the load should be able to be sensitively manipulated upwards and downwards, so that during the transition from upward to downward movement and vice versa the pressure oil inflow on the rotary piston engine has to be switched from primary to secondary. In this phase of change, the rotary engine has no torque because the pressure drops to zero.
  • the spring-loaded work brake 50 takes over the holding torque at this moment and must therefore be designed so large that it can take over the maximum torque of the rotary piston engine.
  • the size and number of springs 78 is to be sized accordingly, as well as the size and number of blades of the working brake 50.
  • a connectable via a separate port 51 to the high-pressure pump high-pressure piston 79 is provided in is able to release the working brake 50, provided that the applied pressure on the high-pressure piston 79 by overcoming the spring forces of the springs 78 is large enough. In practice, it has been proven that this pressure must be between 8 and 12 bar, so that the load does not drop until the required support pressure is built up on the rotary piston engine.
  • Wet-running multi-disc brakes have a particular advantage because they can be connected to the oil cooling system of the entire system through the oil passage. In addition, they are largely abrasion-free, so that the oil pollution is low.
  • the disadvantage is that with oil-filled brake considerable, oil viscosity caused, loss-producing slip performance. According to the Newtonian law of shear stress in an oil gap, the
  • the rotary valve 3 facing annular surface with the pressure windows 33a of the control plate 22 is relatively narrow (smaller diameter difference of the sealing webs). Accordingly, then the difference in the diameter of the mating ring surface between the rotary valve 3 and the axial balance piston 65 is smaller.
  • FIG. 11 shows a longitudinal section through the advantageous development of the rotary piston engine according to the invention.
  • these relationships are clearly shown.
  • the central web diameter 95 and 96 of the control plate 22 see Figs. 11 and 12, and the corresponding central web diameters 97 and 98 of the rotary valve 3, see Figs. 11, 13 and 14, which are shown by the dotted lines, remains first the outer middle Steg tellmesser 99 between the rotary valve 3 and the axial balancing piston 65, see FIGS. 11 and 13, the same, because this together with the web diameter 97 causes the force balance on the rotary valve 3 in the event that the high pressure is supplied to the first annular space 56.
  • the high pressure is supplied to the second annular space 58
  • the axial balance of the rotary valve 3 the new, in the inner diameter annular surface is responsible, which is determined by the new average web diameter 100 and 101.
  • the relief groove 102 and its communication hole 103 may be attached both in the rotary valve 3 and in the axial balance piston 65.

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  • Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Chemical & Material Sciences (AREA)
  • Combustion & Propulsion (AREA)
  • Hydraulic Motors (AREA)
  • Retarders (AREA)

Abstract

L'invention concerne un moteur à piston rotatif hydrostatique, à régime lent, comprenant une section de puissance (1) qui sert de sortie et comporte un stator (4) fixe central, un piston rotatif (6) servant de rotor, et un arbre (2) monté de manière centrale par l'intermédiaire de roulements (10, 11) placés des deux côtés de manière directement adjacente à la section de puissance (1). La commande de l'alimentation des chambres, formées entre les dentures, en fluide de travail et de l'évacuation de ce dernier hors des chambres s'effectue au moyen d'une soupape rotative (3) en forme de disque, qui, selon l'invention, est montée de façon à se déplacer de manière centrale par rapport à l'arbre (2) et au stator (4). Un engrenage est placé entre une denture extérieure d'arbre (14) et une denture intérieure (17) d'une couronne à denture intérieure (28; 92) fixe pour assurer l'entraînement synchrone de la soupape rotative (3). Selon l'invention, l'engrenage est placé exclusivement dans la zone où se trouve l'huile de fuite du moteur et est formé par un engrenage planétaire (80) ou de préférence par un engrenage excentrique.
EP05761659.1A 2004-07-22 2005-07-12 Moteur a piston rotatif hydrostatique Active EP1776525B1 (fr)

Applications Claiming Priority (2)

Application Number Priority Date Filing Date Title
CH01239/04A CH701073B1 (de) 2004-07-22 2004-07-22 Hydrostatischer Kreiskolbenmotor.
PCT/EP2005/007543 WO2006010471A1 (fr) 2004-07-22 2005-07-12 Moteur a piston rotatif hydrostatique

Publications (2)

Publication Number Publication Date
EP1776525A1 true EP1776525A1 (fr) 2007-04-25
EP1776525B1 EP1776525B1 (fr) 2013-08-28

Family

ID=34972717

Family Applications (1)

Application Number Title Priority Date Filing Date
EP05761659.1A Active EP1776525B1 (fr) 2004-07-22 2005-07-12 Moteur a piston rotatif hydrostatique

Country Status (4)

Country Link
US (1) US7832996B2 (fr)
EP (1) EP1776525B1 (fr)
CH (1) CH701073B1 (fr)
WO (1) WO2006010471A1 (fr)

Cited By (3)

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Publication number Priority date Publication date Assignee Title
EP3441613A1 (fr) 2017-08-07 2019-02-13 Siegfried A. Eisenmann Machine hydrostatique à pistons rotatifs et à engrenage
DE202019001218U1 (de) 2019-03-13 2019-04-16 Siegfried Alexander Eisenmann Drehventilantrieb für Zahnrad-Kreiskolbenmotoren
US10421481B2 (en) 2015-09-07 2019-09-24 Volkswagen Aktiengesellschaft Utility vehicle steering system

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US8616528B2 (en) * 2009-01-15 2013-12-31 Parker Hannifin Corporation Integrated hydraulic motor and winch
EP2585719A2 (fr) 2010-06-23 2013-05-01 Siegfried A. Eisenmann Machine hydrostatique à piston rotatif à volume variable en continu
EP2614274A1 (fr) 2010-09-06 2013-07-17 Siegfried A. Eisenmann Entraînement hydrostatique pour un véhicule automobile
EP2607691A1 (fr) 2011-12-22 2013-06-26 Siegfried A. Eisenmann Eolienne avec une pompe hydraulique
DE102011122027B3 (de) * 2011-12-22 2013-04-11 Böhm + Wiedemann Feinmechanik AG Hydrostatischer Kreiskolbenmotor
EP2607683A2 (fr) 2011-12-22 2013-06-26 Böhm+Wiedemann AG Moteur hydrostatique à pistons rotatifs
JP5860695B2 (ja) * 2011-12-28 2016-02-16 Ｋｙｂ株式会社 電動オイルポンプ
JP5767996B2 (ja) * 2012-03-29 2015-08-26 カヤバ工業株式会社 流体圧駆動ユニット
JP5934543B2 (ja) * 2012-03-29 2016-06-15 Ｋｙｂ株式会社 流体圧駆動ユニット
CN102828895B (zh) * 2012-09-07 2015-10-21 镇江大力液压马达股份有限公司 径向支撑轴阀配流摆线液压马达
CN103016336B (zh) * 2012-12-12 2015-01-07 北京动力机械研究所 一种基于行星摆线转子泵的永磁同步电动计量泵
JP6133234B2 (ja) * 2013-07-08 2017-05-24 本田技研工業株式会社 オイルポンプの取り付け構造
DE102013111098B3 (de) 2013-10-08 2014-11-13 4-QM hydraulics GmbH Strömungsmaschine
GB2525704B (en) * 2014-02-14 2016-04-27 Pattakos Manousos Disk rotary valve having opposed acting fronts
DE202014006761U1 (de) 2014-08-22 2015-11-24 Siegfried Eisenmann Hydrostatische Kreiskolbenmaschine nach dem Orbitprinzip
CN106438189A (zh) * 2016-07-09 2017-02-22 镇江大力液压马达股份有限公司 一种超微型摆线液压马达
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US10421481B2 (en) 2015-09-07 2019-09-24 Volkswagen Aktiengesellschaft Utility vehicle steering system
EP3441613A1 (fr) 2017-08-07 2019-02-13 Siegfried A. Eisenmann Machine hydrostatique à pistons rotatifs et à engrenage
DE202019001218U1 (de) 2019-03-13 2019-04-16 Siegfried Alexander Eisenmann Drehventilantrieb für Zahnrad-Kreiskolbenmotoren

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US20080003124A1 (en) 2008-01-03
US7832996B2 (en) 2010-11-16
WO2006010471A1 (fr) 2006-02-02
CH701073B1 (de) 2010-11-30
EP1776525B1 (fr) 2013-08-28

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