WO1997038245A1 - Dispositif hydrostatique de transmission de force a reglage continu - Google Patents

Dispositif hydrostatique de transmission de force a reglage continu Download PDF

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Publication number
WO1997038245A1
WO1997038245A1 PCT/EP1997/001813 EP9701813W WO9738245A1 WO 1997038245 A1 WO1997038245 A1 WO 1997038245A1 EP 9701813 W EP9701813 W EP 9701813W WO 9738245 A1 WO9738245 A1 WO 9738245A1
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Prior art keywords
transmission device
control piston
power transmission
piston unit
anspmch
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Ceased
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PCT/EP1997/001813
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German (de)
English (en)
Inventor
Georg Albersinger
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Individual
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Individual
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Publication of WO1997038245A1 publication Critical patent/WO1997038245A1/fr
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Classifications

    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16H—GEARING
    • F16H39/00—Rotary fluid gearing using pumps and motors of the volumetric type, i.e. passing a predetermined volume of fluid per revolution
    • F16H39/04—Rotary fluid gearing using pumps and motors of the volumetric type, i.e. passing a predetermined volume of fluid per revolution with liquid motor and pump combined in one unit
    • F16H39/06—Rotary fluid gearing using pumps and motors of the volumetric type, i.e. passing a predetermined volume of fluid per revolution with liquid motor and pump combined in one unit pump and motor being of the same type
    • F16H39/26—Rotary fluid gearing using pumps and motors of the volumetric type, i.e. passing a predetermined volume of fluid per revolution with liquid motor and pump combined in one unit pump and motor being of the same type with liquid chambers not shaped as bodies of revolution or shaped as bodies of revolution eccentric to the main axis of the gearing
    • F16H39/30—Rotary fluid gearing using pumps and motors of the volumetric type, i.e. passing a predetermined volume of fluid per revolution with liquid motor and pump combined in one unit pump and motor being of the same type with liquid chambers not shaped as bodies of revolution or shaped as bodies of revolution eccentric to the main axis of the gearing with liquid chambers formed in stationary members
    • F16H39/32—Rotary fluid gearing using pumps and motors of the volumetric type, i.e. passing a predetermined volume of fluid per revolution with liquid motor and pump combined in one unit pump and motor being of the same type with liquid chambers not shaped as bodies of revolution or shaped as bodies of revolution eccentric to the main axis of the gearing with liquid chambers formed in stationary members with sliding vanes carried by the rotor

Definitions

  • the invention relates to a hydrostatic, continuously adjustable power transmission device according to the preamble of claim 1.
  • a power transmission device is known from DE-A-3 143 645.
  • the object of the invention is to provide a hydrostatic, continuously adjustable power transmission device of the type mentioned at the outset, which has a high degree of efficiency and a significantly reduced size with a correspondingly lower weight.
  • the invention is based on the consideration of combining two hydrostatic displacement units not only functionally, but also in a direct structural effect.
  • the constructive approach consists in the spherical shape as the envelope surface of both displacement units, which work according to the vane principle.
  • a largely loss-free, stepless conversion of the power supplied in the form of a speed-torque ratio into any other speed-torque ratio with a high power density can be achieved.
  • the high level of efficiency results from very low specific friction losses, as well as hardly any flow losses due to the direct and throttle-free transfer of the working medium between the drive units and through the same direction of rotation of drive and output.
  • the high power density results from a large volume of the work spaces compared to the small size and from the high achievable speed level.
  • the stepless power transmission device In comparison to known automatic gearboxes, as are currently used in motor vehicle construction, the stepless power transmission device according to the invention requires only a fraction of the construction volume, weight and manufacturing costs.
  • the stepless adjustability of the transmission ratio ensures that, for example, the driving internal combustion engine can always be operated with an optimal speed-torque ratio, which leads to a considerable reduction in fuel consumption.
  • Fig. 1 is a perspective, partially cutaway view of a first
  • FIG. 2 shows an exploded view of the displacer and control piston units of the continuously variable transmission according to FIG. 1, which are accommodated in a spherical envelope surface, the control piston unit being adjustable via an internal mechanism;
  • Fig. 3 is a perspective, partially cutaway view of a
  • Fig. 4 is a perspective, partially cutaway view of a second
  • Embodiment of the power transmission device according to the invention in the form of a continuously variable transmission which in contrast to the first embodiment according to FIGS. 1 to 3 has an adjustment mechanism for the control piston unit which extends over the outside of the displacement units;
  • Fig. 5 is a perspective view of a control piston unit for
  • FIG. 6 to 8 are schematic views of the first embodiment of FIG. 1 with different tilt positions of the control piston unit, namely in the idle position (FIG. 6), in a middle position with the speed ratio 2: 1 (FIG. 7) and an end position with the speed ratio 1: 1 (Fig. 8);
  • Fig. 10 is a schematic development of the annular lateral surface of a
  • Fig. 11 is a perspective, partially cutaway view of a third
  • Fig. 12 is an axial longitudinal section through the third embodiment
  • Fig. 13 is a perspective, partially cutaway view of a
  • FIG. 14 is a partially sectioned axial view of the embodiment of FIG. 13;
  • Control piston unit with the two adjoining displacement units as used in the exemplary embodiments according to FIGS. 11 to
  • FIG. 16 is a partially cutaway illustration of the assembly according to FIG.
  • Fig. 17 is a perspective, partially cutaway view of a fourth
  • Embodiment of the power transmission according to the invention device in the form of a continuously variable transmission which, in contrast to the previous exemplary embodiments, has a control piston unit consisting of two control pistons arranged at an axial distance from one another;
  • Fig. 18 is an axial view of the fourth embodiment shown in Fig. 17;
  • 19 is a perspective, partially cutaway view of a fifth
  • Embodiment of the power transmission device according to the invention in the form of a continuously variable transmission which, in contrast to the previous exemplary embodiments, has a control piston unit consisting of a two-part control piston which can be adjusted on a common pivot axis and has a doubled number of work spaces and volume cycles;
  • Fig. 20 is an axial view of the fifth embodiment shown in Fig. 19;
  • Fig. 21 is a perspective, partially cutaway view of a sixth
  • Embodiment of the power transmission device according to the invention in the form of a continuously adjustable, in particular usable as a differential gear, which has spring-loaded, slidably mounted separating elements with a relief groove on one side, and
  • FIG. 22 is a partially sectioned view of a control piston unit with pressure-dependent throttling of the volume flow between the work spaces for the special use of the control piston unit in transmissions according to the invention, which are operated as differential gears.
  • a first embodiment of a power transmission device according to the invention is shown in perspective in FIG. 1 and in an exploded view in FIG. 2. The associated start, end and middle positions are illustrated schematically in FIGS. 6, 7, 8. Furthermore, FIGS. 9 and 10 schematically show the basic principle of the power conversion as well as the flow and pressure conditions.
  • a housing 1 has a spherical inner surface lc, which serves as an enveloping surface for two spherical cap-shaped displacement units 4, 5, which are arranged on a common longitudinal axis 9a, 10a and with their end faces 4b, 5b (FIG. 6) lie opposite one another at an axial distance .
  • the drive-side displacement unit 4 is rotatably mounted about the longitudinal axis 9a, 10a to the housing section la in a rotary bearing 9b, while the other displacement unit 5 is fixedly connected to the housing section 1b.
  • the space remaining between the end faces of the displacement units 4, 5 is rotationally symmetrical with respect to the longitudinal axis 9a, 10a and is divided into two spaces by means of a control piston unit 6.
  • the control piston unit 6 is rotatably mounted about the longitudinal axis 9a, 10a and takes the output shaft 10 with it.
  • the control piston unit 6 is constructed around a spherical center 6f, which separates and delimits the spaces on both sides of the control piston unit 6 towards the center thereof. This results in two ring-shaped working spaces 2, 3 (FIG. 3) between the housing inner surface 1c, displacement units 4, 5 and control piston unit 6, which are filled with a hydrostatic medium.
  • Each of the two displacement units 4, 5 consists of a spherical cap-shaped inner shell 4a, 5a, which is provided on its end faces 4b, 5b (FIG. 6) with radial slots 4c, 5c (FIG. 6) arranged in a star shape.
  • Wing-like or slide-like separating elements 7 are received in these radial slots 4c, 5c in such a way that the separating elements can be pivoted about the center of their spherical cap-shaped inner shell 4a, 5a.
  • the degree of freedom of this mobility is limited by the spherical inner surface 1c of the housing and the spherical center 6f of the control piston unit 6.
  • This pivoting movement is controlled in that the separating elements 7 on the facing end face 6c, 6d of the control piston unit 6 forcibly.
  • This compulsory system is ensured in that a driver lug 7a is integrally formed on the lower, free end of each separating element 7 and hooks into an annular guide groove 8 at the transition between the control piston unit 6 and the spherical center 6f.
  • the two ring-shaped work spaces 2, 3 are divided into a plurality of work space cells by means of these separating elements 7.
  • an annular groove 5d is provided on the fixed spherical cap-shaped inner shell 5a, which distributes the medium to corresponding bores 5e, which are provided with an integrated check valve 5f.
  • the connection of the annular groove 5d to the outside is made through a hole ld in the housing section lb.
  • the inner shell 4 on the drive side likewise has an annular groove 4d (FIG. 6) which is not visible in FIG. 1 and which corresponds to the annular groove 5d.
  • the annular grooves 4d, 5d serve in both inner shells 4a, 5a to compensate for the changes in volume of the free spaces within the radial slots 4c, 5c (FIG. 6) behind the moving separating elements 7.
  • the mode of operation and the interaction of the two hydrostatic displacement units 4, 5 is determined by the design of the control piston unit 6, which is explained in detail below and which, in the transmission according to FIG. 1, functions as the output, the control of the working medium and the adjustment of the Funding takes over.
  • the main feature of the control piston unit 6 is its end-side annular surfaces 6c, 6d (FIG. 9) which delimit the respective working space 2, 3 and which are arranged non-parallel, ie with mutual inclination to one another.
  • An imaginary cross-sectional plane 6g (FIG. 9) runs through the longitudinal axis 9a, 10a and cuts through the control piston unit 6 such that the part of the cross-sectional area lying above the longitudinal axis 9a, 10a (hereinafter referred to as "upper cross-sectional area A,”) is maximum, and the part lying below the longitudinal axis 9a, 10a (hereinafter referred to as "lower cross-sectional area A 2 ”) is minimal.
  • the upper and lower apex regions are the upper and lower cross-sectional areas A 1 , A 2 corresponding areas of the respective end-side annular surface 6c, 6d of the control piston unit 6 understood.
  • the two opposite pressure halves 2a, 3a like the two opposite suction halves 2b, 3b, are in direct fluidic connection between the upper and lower apex regions through pressure-side or suction-side, slit-shaped control openings 6e.
  • the differential force ⁇ F Via the distance r from this center of gravity to the longitudinal axis 9a, 10a, the differential force ⁇ F generates a torque in a direction of rotation that is identical to the direction of rotation of the drive, with which the desired power transmission to the output shaft 10 (FIG. 2) is effected.
  • control piston unit 6 Another function of the control piston unit 6 is to continuously adjust the speed ratio between the drive shaft 9 and the output shaft 10.
  • the control piston unit 6 is mounted such that it can be tilted about a pivot axis 6a perpendicular to the longitudinal axis 9a, 10a and perpendicular to the cross-sectional plane 6g (FIG. 9).
  • the disk-shaped control piston unit 6 with the cylindrical bore 6i (FIG. 9) of its spherical center 6f is mounted on a spherically shaped hub 10 of the output shaft 10, so that the cylindrical bore 6i is located on the spherical surface of the Hub lOe can roll around the swivel axis 6a.
  • the pivot axis 6a is used for an actuating device guided in the hollow output shaft 10.
  • An actuator 14c (FIG. 1) transmits a static linear movement via a rotary inlet 14 to a drive ring 14a which rotates with the output shaft 10.
  • the driver ring 14a engages via a radial driver pin 14b (which can be axially displaced in the axial longitudinal slots 10d of the output shaft 10) in a push rod 12 which is axially displaceably mounted in the interior of the output shaft 10. In this way, the linear adjustment movement for the control piston unit 6 is transmitted from the actuator 14c into the interior of the output shaft 10.
  • tappets 13 are moved on the push rod 12 by wedge-shaped inclined surfaces 12a which run against one another and which are mounted in a radially movable manner in the spherical center 6f.
  • the upper ends of the plungers 13 lie against the inner bore 6i (FIG. 9) of the center 6f of the control piston unit 6 and, depending on their radial position (which depends on the axial position of the push rod 12), tilt the control piston unit 6 clockwise or counterclockwise around the pivot axis 6a.
  • the control piston unit 6 can be continuously tilted back and forth between the end positions shown in FIGS.
  • FIG. 7 shows a central position in which the angles between the respective end face 6c, 6d and the longitudinal axis 9a, 10a are the same, the speed ratio 2: 1 being set in this position. 6 is the idle position, while in the end position according to FIG. 8 there is a transmission ratio of 1: 1.
  • a bore 6k coaxial to the pivot axis 6a is provided in the spherical center 6f in FIG. 9, into which a driver pin 6h (FIG. 11) is inserted.
  • the rotary movement 9c initiated on the drive side is shown as a linear movement of the corresponding inner shell 4a. If the drive-side inner shell 4a is moved in the direction of the arrow 4e and the pressure half 2a of the work space 2 is viewed, the volume of the individual work space cells of the work space 2 is reduced, as a result of which the medium flows into the pressure half 3a of the opposite work space 3. In order to avoid the increasing medium volume, one would expect the inner shell 5a to move in a direction opposite to the drive-side inner shell 4a.
  • the inner shell 5a is firmly connected to the stationary housing section 1b, only the rotatably mounted control piston unit 6 can evade the increasing media volume, with the result of a rotary movement of the control piston unit 6 in the same direction of rotation as the drive-side inner shell 4a.
  • the volume of the working space cells of the working space 3 increases, as a result of which the medium can be taken up.
  • the medium is transported to the suction half 3b of the work space 3, where the same process takes place analogously to the pressure half 2a of the work space 2 plays, whereby here the volume flow takes place from the fixed inner shell 5a to the driven inner shell 4a.
  • the medium is in turn conveyed in the direction of the pressure half 2a of the working space 2, so that the circuit of the medium closes.
  • the speed ratio n, / n 2 is set by the corresponding setting of the tilting angle of the control piston unit 6. If one considers, for example, the pressure halves 2a, 3a, the drive-side displacement unit 4 delivers a certain volume V per revolution. Analogously, the pressure half 3a of the fixed displacement unit 5 takes up a certain volume V 2 per revolution. The size of the delivered or absorbed volumes per revolution depend on the angle of the respective end face 6c, 6d of the control piston unit 6 with respect to the longitudinal axis 9a, 10a.
  • variable n here means the relative speed of the drive-side displacement unit 4 with respect to the control piston unit 6, while the variable r ⁇ means the relative speed of the control piston unit 6 with respect to the fixed displacement unit 5. Since in this exemplary embodiment the control piston unit 6 represents the output, the actual drive speed determines
  • the second exemplary embodiment shown in FIG. 3 improves by using a hemispherical support shell 11 which surrounds the drive-side displacement unit 4 and is connected to it in a rotationally fixed manner.
  • This displacer unit 4 together with the support shell 11 is slidably or roller-mounted relative to the housing la, as a result of which the sliding friction between the drive-side inner shell 4a and the housing section la of the exemplary embodiment according to FIG. 1 is eliminated.
  • the sealing of the working space 2 is reduced to the impact between the support shell 11 and the housing section 1b.
  • the housing sections 1 a, 1 b are connected in a rotationally fixed manner to the control piston unit 6 and follow their rotation and pivoting movement.
  • This design enables the use of an external adjustment mechanism for the control piston unit 6 with a bow-shaped push rod 12 and rotary lead-in 14.
  • the push rod 12 connects the housing sections 1a, 1b to the rotary lead-in 14 in a moving manner.
  • the output-side displacement unit 4 is non-rotatably coupled to an outer protective housing 100 by means of a hub 5g.
  • this design enables a guide groove 8, which is incorporated in the housing sections 1 a, 1 b, for the forced guidance of the separating elements 7.
  • FIGS. 11 and 12 show an exemplary embodiment of a force transmission device in which, instead of individually movable separating elements 7, wing-like separating elements 7 which are rigidly arranged on a closed carrier ring 71 are used. These wing rings 71, with their smooth surface 71c facing the control piston unit 6, each slide against the respectively adjacent end surface 6c, 6d of the control piston unit 6. Between the separating elements 7, passage openings 71b (FIG. 15) for the working medium are made in the annular surface, which openings are aligned with the slot-shaped control openings 6e of the control piston unit 6.
  • the separating elements 7 dip into radial slots of frustoconical recesses 18 of the inner shells 4a, 5a, which are dimensioned such that the separating elements 7 have a deflection in the recesses 18 (FIG. 15) of the inner shells 4a, 5a corresponding to the tilting position of the control piston unit 6 can execute.
  • frustoconical sealing and compensating rings 18a are embedded in the recesses 18, in the radial slots of which the separating elements 7 slide in.
  • the movement-dependent volume compensation within the sealing and compensation rings 18a takes place via channels 7b (FIG. 16) which extend within the separating elements 7 in the direction of the control piston unit 6.
  • the advantages of the construction according to FIG. 11 are a smaller number of individual parts to be sealed, a loss of individual control of the separating elements 7, and a more favorable, improved sliding friction behavior between the separating elements 7 and sealing and compensating rings 18a compared to the construction according to FIG. 1 11 is the embodiment of the internal adjustment in connection with a spherical center, which consists of a three-part inner ball 61, on which the control piston unit 6 is pivotably mounted.
  • the inner ball 61 comprises a central part 61c, which is connected to the output shaft 10 in a rotationally fixed manner, and two side parts 61a, 61b which are rotatably mounted on the output shaft 10 via roller bearings 61d with respect to the central part 61c.
  • the roller-supported side parts 61a, 61b of the inner ball 61 serve for low-friction and stable support of the inner shells 4a, 5a, which enables better guidance of the inner shells 4a, 5a and improved sealing of the work clearances to the longitudinal axis 9a, 10a.
  • the adjustment of the control piston unit 6 is carried out by means of a transmission lever 17 integrated in the middle part 61c of the inner ball 61, which is actuated by a push rod 12 axially displaceably arranged in the center of the output shaft 10 by means of a rotary introduction 14 similar to that shown in FIG.
  • the transmission lever 17 has ball heads at its ends, which engage in an articulated manner in a blind bore 12b of the push rod 12 and in a radial bore 6m of the control piston unit 6.
  • the spherically thickened shaft between the ball heads of the transmission lever 17 rolls on the inner surface of a radial bore in the central part 61c of the inner ball 61 and in this way transmits the pushing force of the push rod 12 acting on the lower ball head to the upper ball head and thus the control piston unit 6.
  • the control piston unit 6 and the central part 61c of the inner ball 61 are coupled to the output shaft 10 by means of the shafts 6h of the pivot axis 6a and form an internal output.
  • the spherical outer shell for the inner shells 4a, 5a consists of two spherical cap-shaped support shells 11 and a one- or two-part middle ring (ring sections 15a, 15b).
  • the output takes place from the control piston unit 6 through the central ring 15a, 15b into the housing 1 by means of the adjusting shaft 14e, the entire housing 1 rotating is executed.
  • the two support shells 11 are non-rotatably connected to the respective displacement units 4, 5, the central ring 15 forming a unit with the housing 1 and thereby allowing the adjustment shaft 14e to be introduced from the outside.
  • An actuating bracket 14d which is coupled to the adjusting shaft 14e, is pivoted with an actuator 14c which also rotates.
  • FIGS. 15 and 16 show a variant of the power transmission device shown in FIG. 11, which differs in that an external adjustment (similar to that in FIGS. 13 and 14) including housing output and multi-part support shell 11, 15a, 15b for Application comes.
  • an external adjustment similar to that in FIGS. 13 and 14
  • the wing ring 71 and the respective spherical cap 71a can be connected to one another in one piece.
  • the two spherical caps 71a are mounted on the control piston unit 6 arranged therebetween in a precise and low-friction rolling or sliding manner in the radial and axial direction.
  • the one-piece connection of wing rings 71 and spherical caps 71a further reduces the sealing surfaces for work spaces 2, 3 and, in particular, enables a hermetic seal in the center of the transmission.
  • the displacement units 4, 5 in contrast to the exemplary embodiments described so far, are not combined to form a ball, but are arranged with their center points at an axial distance from one another.
  • the inner shells 4a, 5a can be enlarged beyond the center of the hemisphere, as a result of which substantially larger swivel angles can be realized for each of the two control pistons 64, 65 of the control piston unit 6.
  • the separate pivot axes 6a, 6b of the control pistons 64, 65 are located in the region of the facing end face 4b, 5b of the inner shells 4a, 5a, which can be enlarged up to beyond the center of the hemisphere, as a result of which the deflection of the separating elements 7 is significant if wing rings 71 are used reduced and thus their sealing simplified. Anywhere selectable axial distance between the displacement units 4, 5, more complex adjustment mechanisms can be installed, for example for an independent adjustment of the two end faces 6c, 6d of the control piston unit 6.
  • separating elements 7 can be used in this embodiment in single execution and without positive control.
  • These separating elements 7 are preferably controlled by spring elements 72 and / or hydrostatically.
  • the separating elements 7 are provided with a groove 7c on one side, which ensures the volume compensation on the rear side in the guide slot 4c, 5c of the inner shell 4a, 5a.
  • This type of control of the separating elements 7 by means of spring elements 72 and the hydrostatic relief of the separating elements 7 can of course also be used in all other exemplary embodiments with individual separating elements 7. The same also applies to the use of the two-part control piston unit in the exemplary embodiments according to FIGS. 1 to 16 and in the differential gear according to FIGS. 21 and 22.
  • FIGS. 21 and 22 show an application of the principle of a power transmission device according to the invention in the form of a continuously variable transmission as a differential transmission. All of the five exemplary embodiments described above can also be used as differential gears, for example for power distribution in vehicle drive axles are used.
  • the control piston unit 6 serves as a drive unit, while both displacement units 4, 5 with their shaft extensions 101, 102 each form an output unit.
  • the separating elements 7 shown in FIG. 21 are pressed against the control piston unit 6 by prestressed spring elements 72, as already shown in FIGS. 19 and 20 is provided. Furthermore, in the separating elements 7, the hydrostatic relief of the rear slot space is also ensured by a groove 7c on one side.
  • the control piston unit 6 shown in FIG. 22 is specially equipped with a slide-type valve block 66 for use as a differential gear.
  • the valve block 66 is used, depending on the pressure of the medium, to at least specifically close the control-side control openings 6e in order to throttle the volume flow between the working spaces 2, 3 (FIG. 10) and thereby reduce the differential effect to the full Achieve lock.
  • an overload safety device in the form of a pressure relief valve 66a is integrated in this valve block 66 (FIG. 22), which is used during peak loads, e.g. when the vehicle starts up in a flash, a bypass between the pressure and suction sides opens.
  • the O output torque can be infinitely distributed as desired by adjusting the tilt angle of the control piston unit 6 to both displacement units 4, 5.
  • the regulation of the locking effect brings considerable advantages, especially in tough off-road or motor sport applications.
  • the overload protection device 66a can be used above all to improve the starting behavior of vehicles.

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  • Engineering & Computer Science (AREA)
  • General Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • Transmission Devices (AREA)
  • Fluid-Pressure Circuits (AREA)
  • Friction Gearing (AREA)
  • Hydraulic Motors (AREA)

Abstract

Afin d'obtenir un dispositif hydrostatique de transmission de force à réglage continu d'une haute efficacité, de dimensions sensiblement réduites et par conséquent d'un poids réduit, il est proposé que deux unités de refoulement (4, 5) soient montées dans un carter. Les unités de refoulement (4, 5) sont constituées de deux coques intérieures (4a, 5a) en forme de calottes sphériques montées sur un axe longitudinal commun (9a, 10a) de sorte que les faces des coques intérieures (4a, 5a) se font face avec un écartement axial. Une unité de piston de commande (6) montée entre les faces frontales des coques intérieures (4a, 5a) délimite dans le sens axial, avec les deux coques intérieures (4a, 5a), deux chambres de travail (2, 3) remplies d'un milieu hydrostatique. L'ensemble piston de commande peut tourner autour de l'axe longitudinal (9a, 10a) et comprend deux faces (6c, 6d) non parallèles dont l'angle d'inclinaison par rapport à l'axe longitudinal (9a, 10) est réglable. Des éléments séparateurs (7) qui font saillie dans les deux chambres de travail (2, 3) subdivisent chaque chambre de travail (2, 3) en une moitié 'pression' (2a et 3a) et en une moitié 'aspiration' (2b et 3b). Les chambres de travail (2, 3) sont en communication fluidique l'une avec l'autre, par l'intermédiaire des orifices de commande (6e) ménagés dans l'ensemble piston de commande (6). Quand la coque intérieure (4a, 5a) ou l'ensemble piston de commande (6) tournent autour de l'axe longitudinal (9a, 10a), il se crée dans le milieu hydrostatique une différence de pression entre la moitié 'pression' (2a, 3a) et la moitié 'aspiration' (2b, 3b) de chaque chambre de travail (2, 3). Cette différence de pression provoque la transmission de force de la partie motrice (4a ou 5a; 6) à la ou aux parties entraînées (6; 4a et/ou 5a), selon la différence entre la superficie d'une première (A1) et d'une deuxième (A2) section de l'ensemble piston de commande. Le rapport entre les vitesses de rotation (n1/n2) de la partie motrice (4a ou 5a; 6) et de la ou des parties entraînées (6; 4a et/ou 5a) peut être réglé en continu par réglage d'au moins un angle d'inclinaison d'une face (6c, 6d) de l'ensemble piston de commande (6).
PCT/EP1997/001813 1996-04-11 1997-04-11 Dispositif hydrostatique de transmission de force a reglage continu Ceased WO1997038245A1 (fr)

Applications Claiming Priority (2)

Application Number Priority Date Filing Date Title
DE1996114381 DE19614381A1 (de) 1996-04-11 1996-04-11 Hydrostatisches Verstellgetriebe
DE19614381.0 1996-04-11

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Cited By (1)

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Publication number Priority date Publication date Assignee Title
AU2006341656B2 (en) * 2006-04-12 2011-06-02 Telefonaktiebolaget Lm Ericsson (Publ) Differentiated network indication

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Publication number Priority date Publication date Assignee Title
US2401376A (en) * 1943-06-18 1946-06-04 Leon M Sherman Torque converter and control system
FR1165835A (fr) * 1955-06-24 1958-10-29 Pompes hydrauliques rotatives, volumétriques, multicellulaires, à double effet et leur application comme convertisseur de couple
US2927430A (en) * 1955-10-05 1960-03-08 Georgia Tech Res Inst Hydraulic transmission
US3570246A (en) * 1969-03-11 1971-03-16 Joe Floyd Briggs Hydraulic torque converter

Patent Citations (4)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
US2401376A (en) * 1943-06-18 1946-06-04 Leon M Sherman Torque converter and control system
FR1165835A (fr) * 1955-06-24 1958-10-29 Pompes hydrauliques rotatives, volumétriques, multicellulaires, à double effet et leur application comme convertisseur de couple
US2927430A (en) * 1955-10-05 1960-03-08 Georgia Tech Res Inst Hydraulic transmission
US3570246A (en) * 1969-03-11 1971-03-16 Joe Floyd Briggs Hydraulic torque converter

Cited By (1)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
AU2006341656B2 (en) * 2006-04-12 2011-06-02 Telefonaktiebolaget Lm Ericsson (Publ) Differentiated network indication

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