WO1997001711A1 - Electrohydraulic control arrangement for a rotary hydraulic motor - Google Patents

Electrohydraulic control arrangement for a rotary hydraulic motor Download PDF

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Publication number
WO1997001711A1
WO1997001711A1 PCT/EP1996/002818 EP9602818W WO9701711A1 WO 1997001711 A1 WO1997001711 A1 WO 1997001711A1 EP 9602818 W EP9602818 W EP 9602818W WO 9701711 A1 WO9701711 A1 WO 9701711A1
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WO
WIPO (PCT)
Prior art keywords
rotor
valve
motor
position
control
Prior art date
Application number
PCT/EP1996/002818
Other languages
German (de)
French (fr)
Inventor
Eckehart Schulze
Original Assignee
Eckehart Schulze
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
Priority to DE19522768.9 priority Critical
Priority to DE1995122768 priority patent/DE19522768A1/en
Application filed by Eckehart Schulze filed Critical Eckehart Schulze
Priority claimed from DE1996501508 external-priority patent/DE59601508D1/en
Publication of WO1997001711A1 publication Critical patent/WO1997001711A1/en

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Classifications

    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04CROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
    • F04C14/00Control of, monitoring of, or safety arrangements for, machines, pumps or pumping installations
    • F04C14/04Control of, monitoring of, or safety arrangements for, machines, pumps or pumping installations specially adapted for reversible machines or pumps
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04CROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
    • F04C2/00Rotary-piston machines or pumps
    • F04C2/08Rotary-piston machines or pumps of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing
    • F04C2/10Rotary-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
    • F04C2/103Rotary-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
    • F04C2/104Rotary-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 having an articulated driving shaft
    • FMECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04CROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT MACHINES FOR LIQUIDS; ROTARY-PISTON, OR OSCILLATING-PISTON, POSITIVE-DISPLACEMENT PUMPS
    • F04C2/00Rotary-piston machines or pumps
    • F04C2/08Rotary-piston machines or pumps of intermeshing-engagement type, i.e. with engagement of co-operating members similar to that of toothed gearing
    • F04C2/10Rotary-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
    • F04C2/103Rotary-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
    • F04C2/105Details concerning timing or distribution valves
    • F04C2/106Spool type distribution valves

Abstract

The invention concerns an electrohydraulic control arrangement (12) for a rotary hydraulic motor which takes the form of a gear-within-gear motor (11) whose rotor (14) performs two superimposed rotational movements, of which one is eccentric to the central longitudinal axis (19) of the stator (17) and the other is a rotation about the central rotor longitudinal axis (16) and in turn thereby performs a rotational movement about the central longitudinal axis of the stator along a circular path. In order to convert the rotational movements of the rotor (14) into a uniaxial rotational movement of the output shaft (36) of the motor (11), a universal joint shaft (34) which couples the output shaft to the rotor is provided. In order to control the drive, a follow-up regulating valve (39) in the form of a rotary slide valve is provided which operates such that a nominal value of the rotor position is selected under the electrical control of a stepping motor (41), the actual value of the position of the rotor (14) being fed back mechanically via a feedback gear (58) which is adjusted without play and takes the form of a hypotrochoid gear. A sensor element for detecting the actual position value takes the form of a valve piston (57) of the follow-up regulating valve (39). The feedback gear (58) couples the valve piston (57) for movement to the rotor of the gear-within-gear motor such that they rotate uniaxially.

Description

Electro-hydraulic control device for a rotary hydraulic motor

description

The invention relates to an electrohydraulic control device for a rotary hydraulic motor, in which the rotor performs two superimposed rotary movements, one of which takes place eccentrically to the central longitudinal axis of the stator, and the other of which rotates around the rotor-fixed, parallel to the longitudinal axis central rotor longitudinal axis of the stator, which in turn executes a circular movement about the central longitudinal axis of the stator along a circular path, and the conversion of the rotary movements of the rotor into a uniaxial rotary movement of the output shaft of the motor by means of a cardan shaft coupling the output shaft to the rotor is carried out, and a position control circuit working with a variable position setpoint is provided, in which a uniaxially rotationally driven mechanical sensor element which is coupled to the rotor without play is provided for detecting the actual position value.

Such a control device is known from the German utility model G 93 08 025 in connection with a hydraulic motor operating according to the gerotor principle.

Such a hydraulic motor has an annular stator, which is provided on its inside with longitudinal grooves which are grouped axially symmetrically with respect to the central longitudinal axis of the stator and are offset from one another by ribs with an arcuate contour. The rotor is in the form of a star-shaped disk, the thickness of which corresponds to that of the stator ring. The star spikes of the rotor have a convex curvature and usually connect with a smooth curvature to flat-concave contour regions, which each run between two radially projecting heads of the rotor. The inner contour of the stator and the outer contour of the rotor are matched to one another in such a way that the rotor, in any of its possible rotational positions which it can assume during a 360 ° rotation, makes line contact with each of the ribs extending in the axial direction of the stator. The multiplicity of the axial symmetry of the stator is 1 higher than the multiplicity (numeracy) of the axial symmetry of the rotor, which in practical cases is at least 4. With such a configuration of the rotor and the stator, the chambers, which are delimited in a pressure-tight manner in the radial and azimuthal direction and by housing plates in the axial direction, the number of which corresponds to that of the grooves of the stator, are in any azimuthal position of the rotor different volumes, which change continuously with a continuous rotary movement of the rotor, so that by valve-controlled pressurization of those chambers which, viewed in the intended direction of rotation of the motor, enlarge and pressure relief of those chambers which change in the same direction of rotation reduce, the rotor is driven in the desired direction of rotation. This type of drive control, which requires an off-axis mounting of the rotor with respect to the central longitudinal axis of the stator, has the consequence that the axis of rotation of the rotor parallel to the central longitudinal axis of the stator has a number of circular movements corresponding to the number of its protruding teeth executes the central longitudinal axis of the rotor when the rotor makes a 360 "revolution, the direction of rotation of this circular movement of the axis of rotation of the rotor being opposite to the rotary movement of the rotor itself.

In order to achieve the freedom from play of the drive coupling of the output shaft of the hydraulic motor, with the rotor carrying out the superimposed rotary movements, which is desired for control-technical reasons, threaded end sections of an articulated shaft, with which they mesh with an internal toothing of the tube and an internal toothing of the output shaft stands, divided and torsionally braced against one another, the threaded end sections having spherically curved teeth in order to be able to compensate for an axial offset which is caused by wobble movements of the cardan shaft. The actual value detection of the position of a part which is rotated by means of the motor, for example, when the known rotary hydraulic motor is used as an actuator, can be carried out in a known manner by means of an electronic or electromechanical rotary position encoder system which encodes ¬ tion of the rotational position of the output shaft in characteristic electrical signals allows for that too A setpoint-actual value comparison of an electronic control unit of the drive can be fed, from whose comparative processing with setpoint-characteristic predetermined signals this electronic control unit generates control signals for the valve control of the engine.

Regardless of the basic suitability of the known control device for precise positioning operation, it nevertheless has at least the following disadvantages:

Since the actual position value is determined by monitoring the azimuthal position of the output shaft, the characteristic frequency of the control system must be significantly reduced compared to the natural frequency of the spring mass system formed by the load and its coupling to the output shaft in the event of a rapidly increasing deviation from the target and actual position, to rule out a too "violent" backlash of the control, which could otherwise lead to an increase in vibrations and, in extreme cases, to damage to the drive train.

In applications where such a hydraulic motor is used as the drive unit of a linear drive, for example a spindle drive, and two such linear drives are provided, for example, to prevent a workpiece or tool from moving along a path curve by superimposing the linear movements in two mutually perpendicular coordinate directions to achieve, because of the low usable control loop gain, a relatively low web speed must be controlled so that the web can be followed with sufficient accuracy, ie acceptable deviations from its ideal course. The consequence of this is an overall low web tracking speed, which is of course undesirable for production reasons.

The object of the invention is therefore to improve a control device of the type mentioned at the outset such that a control circuit provided for drive control of the hydraulic motor can be operated with high circuit gain and nevertheless the risk of damage to the drive train is largely ruled out .

This object is achieved according to the invention in that the sensor element used for the actual position value detection with the rotor by means of its own, in addition to the drive train formed by the rotor, the cardan shaft and the output shaft, via which a predominant part of the engine torque is slid as the useful torque, the provided sensor drive train, which is free of play, is motion-coupled.

According to this, the it position, which can be detected as azimuthal deflection - rotation - of the rotor relative to a reference plane, which contains the central, housing-fixed axis of rotation of the rotor, for comparison with a mean setpoint input drive, which is electrical is controllable. The advantageous consequence of this is that the "softness" of the drive train on the output side, in particular a torsional deformation of the cardan shaft that couples the rotor to the output shaft, depends on the actual value information, however it is obtained, for example by means of an electronic ¬ or electromechanical rotary position sensors, can not have an effect, so that the position control loop used for position control of the rotor has a high degree of rigidity and, accordingly, the circular gain K v dieεeε control loop, which is generally due to the relationship

Figure imgf000008_0001

given that C is the stiffness and m is the mass of the mass-spring system of the control circuit, can be correspondingly high, the stiffness c being essentially due to the - low - compressibility of the position Control loop existing 01 column is given and the mass m is essentially determined by the mass of the rotor.

By regulating the rotor position with high circular gain, rapid control of the position of the element driven by the hydraulic motor is achieved.

One to implement the superimposed rotational movements The gear unit suitable for monitoring the uniaxial rotary movements of the mechanical sensor element of the position actual value detection device in a simple manner and which can also be sufficiently implemented with simple means to meet the requirement of freedom from play, is a trochoidal gear unit with low axial and radial Dimensions can be realized.

In this case, a configuration as a hypotrochoid gear is preferred, which consists of a ring gear with internal toothing and a pinion in meshing engagement with its toothing, whose pitch circle diameter d is smaller than the pitch circle diameter D of the ring gear toothing, the difference amount e corresponds to the diameter of the circle on which the rotor-stationary axis of rotation of the rotor rotates about the stator-stationary axis of rotation of the motor. If the ring gear is provided in a coaxial arrangement of its internal toothing with the rotor-fixed axis of rotation and the pinion in a coaxial arrangement with the stator-fixed central axis is provided with the mechanical sensor element of the position actual value detection device, this results in the same direction of rotation of the rotor and above that Hypotrochoid gear driven actual value detection element, while with the rotationally fixed arrangement of the pinion on the rotor and arrangement of the ring gear on the driven mechanical sensor element, the position actual value detection device results in the opposite direction of rotation of the rotor and the actual value detection element.

If this is uniaxially rotatory via the further gear Drivable feedback element Functional element of the mechanical actual value feedback device of a follow-up control valve provided for the movement and position control of the rotary hydraulic motor is that it works with an electrically controllable setpoint specification, so it is particularly advantageous if this follow-up operation Control valve is designed as a rotary slide valve with rotatably drivable piston and housing elements, the azimuthal deflection of which is limited against one another.

Both in the configuration of the follow-up control valve, in which the centrally arranged valve piston is provided with a pinion, which is in meshing engagement with the ring gear of the rotor, and the housing of the follow-up control valve by means of the setpoint specification motor deε Valve can be driven, as well as in that in which the centrally arranged valve piston can be driven by means of the setpoint specification motor and the valve housing is provided with the ring gear meshing with the pinion of the rotor of the hydraulic motor, the follow-up control valve is designed in a simple manner realizable by the fact that a housing is designed as a valve bushing, which is rotatably arranged in a centrally through bore of a valve housing block with a sealing sliding fit that is firmly connected to the engine housing. The required freedom from play of the feedback drive can then be realized in a simple manner by providing a valve spring arrangement which permanently acts between the valve bushing and the piston of the follower control valve Torque generated which is smaller than the holding torque of the setpoint specification motor in its de-energized state and also smaller than the holding torque of the rotary hydraulic motor when the pressure supply is switched off.

Further details of the invention result from the following description of exemplary embodiments with reference to the drawing. Show it:

1 shows an electrohydraulic drive unit with a gerotor-hydraulic motor as a power drive and a follow-up control valve which is electrically setpoint-controlled and designed as a rotary slide valve by means of a stepper motor and which works with mechanical position / value feedback via a feedback gear Simplified longitudinal section table,

FIGS. 1 a and 1 b show further functional positions of the run-on control valve according to FIG. 1 in the corresponding representation;

Fig. Lc is a diagram for explaining the function of

Follow-up control valve of the drive unit according to FIG. 1,

2 shows the gerotor motor of the drive unit according to FIG. 1, in section along the line II-II of FIG. 1, 3 shows details of the overrun control valve which can be used in the drive unit according to FIG. 1 in a longitudinal section representation corresponding to FIG. 1, FIG.

3a shows a section along line IIIa-IIIa of FIG. 3,

3b shows a section along line IIIb-IIIb, in each case with the central position of the follow-up control valve associated with the shutdown of the hydraulic motor of the drive unit,

FIG. 3d the sectional representations corresponding to the representations of FIGS. 3b and 3c and FIG. 3e corresponding to the explanation of one of the two flow positions of the follow-up control valve of the drive unit according to FIG. 1

3f, FIGS. 3d and 3e, corresponding sectional representations and 3g positions to explain the second flow position of the follow-up control valve and

3h shows a section along line III-IIIh of FIG

3 to explain the backlash of the feedback gear of the control device according to FIG. 3 mediating tensioning device.

The electro-hydraulic drive unit, designated overall by 10 in FIG. drove a rotary hydraulic motor 11 designed as a gerotor as well as an electro-hydraulic control device, designated overall by 12, which is combined with the hydraulic motor 11 to form a compact structural unit 13.

The gerotor motor 11, for which explanation is also referred to FIG. 2, has a star-shaped rotor 14, which has a multiple number of rotations with respect to a central rotor axis 16, and four-fold rotational symmetry in the special embodiment shown. as well as an annular stator 17 surrounding the rotor 16, which has the basic shape of a toothed ring with an internal toothing which forms the radially outer boundary of a stator interior 18 receiving the rotor 14, which in turn relates to the central longitudinal axis 19 of the motor 11 and the drive unit 10 as a whole is designed to be rotationally symmetrical, the multiplicity of the stator symmetry being 1 higher than that of the rotor 14 and thus 5-fold in the special embodiment shown.

The radially inwardly projecting teeth 21 of the stator 17 are formed as convex ribs extending parallel to the central longitudinal axis 19 with circumferential surfaces 22 curved in a circular arc.

The teeth 23, ie the radially most protruding areas of the star-shaped rotor 14, are also convexly curved and close with a radius of curvature that is smaller than that of the ribs 21 of the stator 17 with a smooth curvature on flat concavely curved jacket regions 24 of the rotor 14 which mediate between the teeth 23 and whose radius of curvature is greater than that of the jacket surfaces 22 of the ribs 21 of the stator 17.

The rotor 14 and the stator 17 of the gerotor 11 have the same axial thickness and are arranged between annular disks 26 and 27 of the motor housing, which is denoted overall by 28 and which form the housing-fixed, axial limitations of the five drive chambers 18 1 to 18 5 of the gerotor 11 which are delimited radially on the outside of the housing by the stator 17 and are movable radially on the inside by the rotor 14, whose alternately convex and concave shell contour profile is matched to that of the stator 17 in such a way that the rotor 14 is in all possible azimuthal positions with respect to one another a reference plane containing a central axis 16, the orientation of which can be arbitrarily selected, is in contact with each of the toothed ribs 21 of the stator along a circumferential line 29 which run parallel to the central longitudinal axes 16 and 19 of the rotor 14 and the stator 17, two ge in the circumferential direction along these contact lines 29 εehen adjacent drive chambers are closely delimited from each other.

By periodically alternating pressurization and relief of the drive chambers 18 to 18 5, the rotor 14 can be controlled to carry out rotations about its central axis 16, which in this case when the rotor 14 is in the time represented by the head 31 rotates clockwise around the central axis 19 of the stator, the number of rotations of the central axis 16 of the rotor around the central axis 19 of the stator 14 compared to the number of revolutions of the rotor 14 around its central Axis 16 around the multiplicity of the rotor symmetry, thus four times in the special embodiment shown, the number of revolutions of the rotor about its central axis 16 is higher than that.

The pressure application and relief of the drive chambers 18, biε 18 5 in the direction of rotation in such a way that pressure is applied to each drive chamber that can expand and those chambers to the pressureless storage container of the pressure supply unit (not shown) To be relaxed, the volume of which decreases with the current direction of rotation of the rotor 14, a rotary slide valve 32 synchronized mechanically with the rotor 14, indicated only schematically in FIG. 1, conveys its piston 33 with the rotor 14 of the gerotor 11 an articulated shaft 34 is rotationally coupled to motion, through which the superimposed, epicycloidal movements of the rotor 14 of the gerotor 11, namely the rotation of the rotor 14 about its central longitudinal axis 16 and its epicycloidal movement about the central longitudinal axis 19 of the stator 17 of the gerotor 11, in a uniaxial rotational movement of the valve piston 33 of the rotary slide valve 32 are converted around the central longitudinal axis 19 of the drive unit 10. By means of this articulated roller 34, the uniaxial-rotary loading Movement of the output shaft 36 of the gerotor 11 is achieved, which is connected in a rotationally fixed manner to the valve piston 33 and, for this purpose, is embodied integrally with the valve piston 33 of the rotary slide valve 32 in a suitable configuration.

The direction of rotation of the rotor 14 is determined by the way via which its control connection 37 or 38 pressurized hydraulic medium is supplied to the gerotor motor 11 and can flow to the unpressurized reservoir of the pressure supply unit, the speed of the rotor 14 being adjusted by the control - The amount of the hydraulic medium flowing through the gerotor 11 in the time unit is controlled.

The control functions required for this are conveyed by the electrohydraulic control device 12, which is designed as a position control circuit for the positioning of the rotor 14 of the gerotor motor 11 and with an electrically controllable position setpoint specification and mechanical position value feedback is working.

The electrohydraulic control device comprises an overrun control valve, designated 39 overall, which conveys the function of a 4/3-way proportional valve, that by actuating an electric stepper motor 41 in alternative directions of rotation in alternative functional positions I (FIG. 1) and II ( 1 a) is controllable, which in turn corresponds to alternative directions of rotation of the gerotor motor 11. In the first 1, the function position I of the wake control valve 39 are the high pressure (P) supply connection 42 with the A control connection 37 of the gerotor motor 11 and the B control connection 38 with the pressureless tank (T). - Supply connection 43 of the follow-up control valve 39 connected, which may correspond to an operating state of the gerotor motor 11 in which its rotor 14, viewed in the direction of the arrow 44 in FIG. 1, ie from the step motor 41, is located rotates in the direction of the central longitudinal axis 19 of the gerotor motor 11 in the clockwise direction represented by the arrow 31 in FIG. 2.

If the stepper motor 41 is driven in the opposite direction of rotation, the overrun control valve 39 reaches its alternative function position II (FIG. 1 a), in which the B control connection 38 of the gerotor motor 11 with the P-supply connection 42 of the follow-up control valve 39 and its A-control connection 37 are connected to the T-supply connection 43 of the follow-up control valve, and thereby the gerotor motor 11, seen in the illustration in FIG. 2, for executing rotary movements in the direction of arrow 46, ie is driven counterclockwise.

The follow-up control valve 39, in accordance with its function as a 4/3-way valve, is designed so that whenever the follow-up control valve has one of its two alternative functional positions I and II, the drive functions of the gerotor motor 11 in alternative directions of rotation are assigned, in which the other direction of rotation function-assigned function position is switched, this switching leads to an intermediate position 0 (FIG. 1b), in which both the A control connection 37 and the B control connection 38 of the gerotor motor 11 against the P supply connection 42 and the T -Verεorgungε- connection 43 of the follow-up control valve 39 are shut off.

The follow-up control valve 39, in accordance with its function as a proportional valve, is designed so that between control positions φ - and Φ z (FIG. 1c) the respective maximum flow cross-section Q max of the flow paths released in the functional positions I and II of the follow-up control valve and the blocking position 0, which corresponds to the intermediate position <J> 2 , the flow cross sections of these flow paths 47 and 48 or 49 and 51 vary between the maximum value ° max and the value ° monotonously.

The follow-up control valve 39 is mounted on the housing 28 thereof on the side thereof remote from the output shaft 36 of the gerotor motor 11. It comprises a housing 52 which is firmly connected to the housing 28 of the gerotor motor 11 and which has a continuous, central bore 53 in the central bore 19, which is coaxial with the central longitudinal axis 19 of the gerotor motor, in which seals against the housing bore 53 , a cylindrical tubular valve bush 54 is rotatably arranged, which can be driven in rotation by means of the stepper motor 41, which in turn - stator-fixed - is mounted on the housing 52 of the overrun control valve. In the continuous central bore 56 of the valve bush 54, sealed against it, a basic valve piston 57 is arranged to be rotatable about the central axis 19 of the gerotor 11, the valve piston 37 being designated by 58 with a - free of play - overall Feedback gear is rotatably coupled to the rotor 14 of the gerotor motor 11.

By means of this feedback gear 58, the valve piston 57 of the follow-up control valve 39 is driven in the same direction of rotation in which the valve bushing for setting the position setpoint 54 is driven by the stepper motor 41.

In a stationary operating state of the control device 12, which corresponds to a constant rate of change of the position setpoint and a constant angular velocity of the rotor 14 around the momentary position of a rotary axis 16, the central valve piston of the cylindrical valve bush 54 races by one - azimuthal - Follow-up error ΔΦ according to which the opening cross section of the flow paths 47 and 48 or 49 and 51 specified in the respective functional position I or II of the follow-up control valve 39 is determined, via which this leads to the gerotor motor 11 and from this pressure medium - hydraulic oil - which is returned to the pressure-less storage container of the pressure supply unit under the operating pressure which is established under load in the for compliance the rotational speed of the rotor 14 can flow required flow rate. If the position setpoint specification, which is effected by actuation of the stepper motor 41 with position setpoint specification pulses, is ended, with the result that the valve bush 54 remains in its position reached up to then, the further rotation of the rotor 14 leads in the previously controlled direction of rotation so that the overrun control valve 39, after its central valve piston 57 has been rotated further by the azimuthal overrun path .DELTA..phi., reaches its blocking position 0 (FIG. 1b) in the equality of the actual position and the desired position of the rotor 14 of the gerotor motor 11 is given, and, if they are rotated beyond the desired position - should overshoot - to the effect that the overrun control valve 39 arrives in the function position assigned to the opposite direction of rotation, with the Consequence that the gerotor motor 11 is automatically activated in the opposite direction of rotation to its original direction of rotation, and dad It finally arrives in the controlled target position as the actual position.

If the drive unit 10 is used to position a work piece or a machine element, this drive coupling with the gerotor motor 11 can be realized by a rack and pinion drive (not shown) which has a toothed rack fixedly connected to the element to be positioned and one with this element This meshing pinion, which is connected to the output shaft 36 of the gerotor motor 11 in a rotationally fixed manner, εo iεt, taking into account the transmission ratio this rack and pinion drive as well as the feedback gear at every moment of operation of the drive unit 10 the position setpoint by the algebraic sum of the stepper motor 41 supplied to the stepper motor 41 at that moment by an electronic control unit 59, by which the stepper motor 51 and so that the gerotor motor 11 can also be controlled in alternative directions of rotation, whereby it is assumed that the rotor of the stepper motor is controlled by each of these control pulses to perform an incremental rotation by the same angular amount ff <f and that for control of the stepper motor 41 in the opposite direction of rotation used output pulses of the electronic control unit 59, which are fed to the stepper motor at different control inputs 61 and 62, are "counted" with the opposite sign, algebraically summed.

If, on the other hand, the drive unit 10 is used as a rotary drive for a workpiece or machine element which is rotatably driven during machining and which is rotationally coupled directly or via a gear to the output shaft 36 of the gerotor motor 11 of the drive, it essentially being based on the Rotation speed arrives, so the relevant speed setpoint is essentially determined, ie again apart from gear ratio ratios, by the frequency with which the drive pulses for the stepper motor 41 are output by the electronic control unit 59. For both operating modes, it is advantageous in terms of a sensitive controllability of the gerotor motor 11 if the incremental change in position of the rotor of the stepper motor 41 and thus also of the valve bushing 54 of the follower control valve 39 associated with each control pulse is small compared to that maximum permissible overrun error Δ (D max is by which the actual position of the central valve piston 57 lags behind the azimuthal position of the valve bush 54, with the position of which the respectively set position setpoint is linked.

In a typical design of the follow-up control valve 39 of the drive unit 10, based on that position (φ 2 ) of the valve bush 54 and the central valve piston 57 relative to one another, in which the follow-up control valve 39 assumes its blocking position 0, the angle by the valve bushing 54 and the valve piston 57 can be rotated relative to one another in the sense of actuating the wake control valve 39 in the sense of taking up their alternative functional positions I and II corresponding to their respective maximum flow cross-section, a typical amount of 30 °, with this actuating range limited by the stop action between the valve bush 54 and the valve piston 57; the incremental angle of rotation X §, on the other hand, by which the rotor of the stepping motor 51 rotates when it is driven with a setpoint input pulse from the electronic control unit 59, has a typical value of 1/10 degree and is therefore approximately 1/300 of the maximum "one-sided" control angle small against them. With this design of the overrun control valve 39, the maximum permissible - azimuthal - overrun error Δ ( ^ maκ , when used, a good response of the control device 12 can still be achieved, a typical amount of 20 °, which corresponds to 2/3 of the maximum possible control angle of the overrun control valve 39 .

The fact that the rotary coupling of the central valve piston 57 of the overrun control valve 39 to the rotor 14 of the gerotor motor 11 provides feedback gears 58, for which explanation reference is also made to FIG. 2, in the case of the one shown in FIG. 1 ¬ th embodiment, a ring gear 63, which is connected in a coaxial arrangement with the common central longitudinal axis 19 of the gerotor motor 11 and the follow-up control valve 12 in a rotationally fixed manner with the central valve piston 57 and arranged on the side facing the gerotor motor 11, as well as a pinion 64 connected to the rotor 14 of the gerotor motor 11 in a rotationally fixed manner, which with its outer toothing 66 is in meshing engagement with the inner toothing 67 of the hollow wheel 63 and is coaxial with respect to the central, rotor-fixed longitudinal axis 16 thereof and therefore with respect to the common central longitudinal axis 19 of the gerotor motor 11 and the follow-up control valve is arranged off-axis.

The number of teeth zl of the pinion is significantly smaller than the number of teeth z2 of the internal toothing 67 of the hollow gear 63 and corresponds approximately to their half value. For the embodiment selected for the explanation, it is assumed that the number of teeth z1 of the pinion is 16 and the number of teeth z2 of the pinion 63 is 30. To explain the function of the feedback transmission 58 designed in this way, the operating state of the gerotor motor 11 is assumed, in which its rotor 14 is, as seen in the direction of the arrow 44 in FIG. 1, in the clockwise direction, ie in the direction of the arrow 31 2, rotates.

Starting from the configuration shown in FIG. 2 with respect to the vertical plane of symmetry 68 of the stator 17 of the gerotor motor 11 overall symmetrical configuration thereof, in which the central longitudinal axis 16 of the rotor 14 below the central longitudinal axis 19 of the stator 17 Gerotor motor 11 runs and the volume of the lowest-lying drive chamber 18 corresponds to the minimum value, the central longitudinal axis 16 of the rotor 14 describes a complete circular path around the central longitudinal axis whenever it rotates 90 ° clockwise 19 of the stator 17 of the gerotor counterclockwise. The pinion 64 thus, while undergoing a 360 ° clockwise rotation, makes four counterclockwise rotations about the central longitudinal axis 19 of the stator 17 of the gerotor motor 11, with the general rule that with N- Numerous symmetry of the rotor 14, the central longitudinal axis N of which rotates around the central longitudinal axis 19 of the stator 17, while the rotor 14 undergoes a 360 "revolution.

The total transmission ratio Ig with which revolutions of the rotor are converted into revolutions of the central valve piston 57 of the overrun control valve 39, iεt in the kinematics described by the relationship

Ig = N - (N + 1) z- / z 2 (1)

given.

For the exemplary embodiment chosen for explanation with 4-fold symmetry of the rotor 14 of the gerotor motor 11 and the ratio z, / z 2 of the number of teeth z and z 2 of the pinion 64 and the meshing ring gear 63 of the central valve piston 57 of 16/30 results in a value of 4/3 for the overall transmission ratio I, the kinematics of the feedback gear 58 leading to the central valve piston 57 rotating in the opposite direction to the rotor 14 of the gerotor motor 11. In order to control a clockwise rotary movement of the output shaft 36 of the gerotor motor 11, the valve bush 54 of the follower control valve 39 must be driven counterclockwise by means of the stepper motor 41 and at an angular speed that is 1/3 higher than the desired rotational speed. output shaft 36.

The embodiment according to FIG. 1d differs from that described with reference to FIG. 1 with regard to the design of the feedback gear 58 'in that the ring gear 63' on the rotor 14 of the gerotor motor 11 and the pinion 64 'on the central valve piston 57 of the follow-up control valve 39 are arranged, which otherwise have the same design with regard to their number of teeth z and z 2 are required.

The feedback gear 58 'according to FIG. 1d, under the same operating conditions as assumed to explain the gear 58 according to FIG. 1, results in a rotational drive of the central valve piston 57 which is the same as the rotation of the rotor 14 of the gerotor motor 11, the now Total translation I 'through the relationship

I g '= N (z 2 / z- - 1) + z 2 / Zl (2)

is given, in which N again denotes the multiplicity of the rotational symmetry of the rotor 14 of the gerotor motor 11, z, the number of teeth of the pinion 64 'and z 2 the number of teeth of the internal toothing of the hollow gear 63'. Accordingly, this transmission ratio I ' g / the same dimensioning of the hollow wheel 63' and of the knurling 64 ', as assumed for the embodiment according to FIG. 1, has the value 5.375.

1 can also be realized in that the ring gear is formed by an engine-side, internally toothed end section of the valve bush 54 and instead of that the central valve piston 57 is driven by the stepper motor 41.

The follow-up control valve 39, which is only schematically illustrated in FIGS. 1 and 1a and 1b to explain its function, has more in detail than that 3 and the cross-sectional representations of FIGS. 3a to 3g, to which reference is made below:

The housing 52 of the follow-up control valve 39 is provided with a total of four annular grooves 71 to 74 open to its central bore 53, with which the radially outer, flat annular grooves 76 to 79 of the essentially tubular valve bushing 54 are in constant communication Connect. The annular grooves 71 to 74 of the housing 52 of the follow-up control valve 39 and the annular grooves 76 to 79 of the valve bushing 54 of the follow-up control valve communicating with them are arranged at the same distance from one another, seen along the central longitudinal axis 19 thereof , the inner groove 71 of the follow-up control valve housing 52 closest to the gerotor motor 11 and the inner groove 74 of the follow-up control valve housing 52 which is furthest away from the gerotor motor 11, each individually, with one of the supply connections 42 and 43 of the follow-up control valve 39 are communicatively connected, as can also be seen from the schematic illustration in FIG. 1.

In the exemplary embodiment chosen for explanation, it is the inner and outer grooves 71 and 76 of the housing 52 closest to the rotor 14 and the valve bush 54 of the follow-up control valve 39 that are in constant communication with the high pressure (P) valve. supply connection 42 of the follow-up control valve 39, while the annular grooves 74 and 79 of the gearbox arranged at the greatest axial distance from the gerotor motor housing 52 and the valve bushing 54 of the follow-up control valve 39 are kept in constantly communicating connection with the T-supply connection 43 of the follow-up control valve, which on the one hand communicates with the unpressurized reservoir of the - not shown - pressure supply unit stands.

The inner groove 72 adjacent to the P-supply groove 71 of the follow-up control valve housing 52 is connected to the A control connection 37 of the gerotor motor 11 in the exemplary embodiment shown via a connection channel (FIG. 1) designated as 81 (FIG. 1) of the valve housing 52 . The inner groove 73 of its housing 52, which is adjacent to the T-supply groove 74 of the housing 52 of the follow-up control valve 39, is via a B connection channel of the valve housing 52 of the follow-up control valve 39, designated overall by 82, with the B control connection 38 of the gerotor motor 11 communicating connected, as can be seen from the schematic illustration in FIG. 1.

The central valve piston 57 is provided with a P-circumferential groove which is coaxial with the P-supply groove 71 and the radially outer P-ring groove 76 of the valve bushing 54 and which has radial P-holes 84 in the valve bushing 54 with its P-ring groove 76 and thus also with the P - Supply groove 71 of the housing 52 is in a constantly communicating connection.

In addition, the central valve piston 57 with a T-supply groove 74 of the housing and the radial The outer T-ring groove 79 of the valve bush 54 is provided with a coaxial T-circumferential groove 86, which in turn is in constant communication via radial transverse bores 87 of the valve bush 54 with its T-ring groove 79 and thus also with the T-supply groove 74 of the housing 52 (Fig. 3b).

From the P-circumferential groove 83 of the central valve piston 54, two P-control grooves extend diametrically opposite one another and extend in the longitudinal direction

88 and 89, which end at an axial distance from the T circumferential groove 86 of the central valve body 57, the axial extent of these P control grooves 88 and

89 is chosen so that - in the axial direction - there is an overlap of these control grooves with the B-ring groove 78 of the valve bush 54 and thus also with the B-ring groove of the valve housing 52.

Also from the T-circumferential groove 86 of the central valve body 57 are two diametrically opposed T-control grooves 91 and 92, which extend in the longitudinal direction to the P-circumferential groove 83 of the central valve body 57 and which are at an axial distance from the P-circumferential groove 83 of the central valve body 57, the axial extent of these T-control grooves 91 and 92 again being chosen such that in the axial direction these T-control grooves 91 and 92 overlap with the A-ring groove 77 of the valve bush 54 and thus also with the A-ring groove 71 of the valve housing 52. The common longitudinal center plane 93 of the P control grooves 88 and 89 and the common longitudinal center plane 94 of the T control Grooves 91 and 92 of the central valve piston 57 run at right angles to one another and cut along the central longitudinal axis 19 of the follow-up control valve 39.

The P control grooves 88 and 89 and the T control grooves 91 and 92, seen in the circumferential direction of the central valve piston 57, have the same azimuthal width CL of 40 °.

The valve bushing 54 is provided with two A-control channels 96 and 97 which are aligned with one another and radially penetrate their sheath and which radially on the outside lie in the A-ring groove 77 which is in constant communication with the A-ring groove 72 of the housing 52 the valve bushing 54 open.

These A-control channels 96 and 97 are each formed by a radial bore 98 with a circular cross-section and by these widening circumferential expansion slots 99 and 101, whose clear width measured in the axial direction is smaller than the diameter of the central, radial bore 98, and its azimuthal depth measured in the circumferential direction is dimensioned such that the azimuthal width measured on the circumference of the central valve piston 57 for these A control channels 96 and 97 corresponds to the azimuthal distance of the P control grooves 88 and 89 corresponds to the T-control grooves 91 and 92, ie has the value of 50 ° in the example chosen for the explanation. The diameter of the central bores 98 of the A control channels 96 and 97 is approximately smaller than the clear width of the P control grooves 88 and 89 and T control grooves 91 and 92 measured in the axial direction, which in turn is slightly smaller than the clear width of the A-ring groove 72 of the housing 52 measured in the axial direction of the follow-up control valve 39.

Furthermore, the valve bushing 54 is provided with two B control channels (FIG. 3c) which radially penetrate the jacket of the valve bushing 94 (FIG. 3c) and which correspond to their design according to the A control channels 96 and 97 and which radially dial into the outside the B-ring groove 72 of the housing 52 in a constantly communicating connection B-ring groove 78 of the valve bush 54 open.

The common central longitudinal axis 104 of the B control channels 102 and 103 of the valve bush 54 is offset by 90 ° in the azimuthal direction from the common central longitudinal axis 106 of the A control channels 96 and 97 (FIG. 3b).

The A-connection channel 81 of the overflow control valve 39 is through a transverse bore 107 running at right angles to its central longitudinal axis 19 and at a radial distance therefrom, which communicates with the A-ring groove 72 of the housing 52 of the overflow control valve 39 in a communicating manner bond is formed, and a longitudinal bore 108 of the valve housing 52, which communicates with the transverse bore 107 and communicates radially outside the A-ring groove 72 of the housing 52, is formed, which with the A control connection 37 of the gerotor motor 11 or of its rotary Directional control valve 32 is in communicating connection.

In an analogous manner, the B connection channel 82 of the overrun control valve 39 is through a transverse bore 109 running at right angles to its central longitudinal axis 19 and at a radial distance therefrom, which bores with the B ring groove 73 of the housing 52 of the overrun control valve 39 in communicating connection is formed, and a longitudinal bore 111 of the valve housing 52 running radially outside the B-ring groove 73 of the housing 52, which communicates with the transverse bore 109, is formed, which with the B control connection 38 of the gerotor motor 11 or 11 . its direction of rotation control valve 32 is in communicating connection.

The cross-bores 107 and 109, each forming a section of the A-connection channel 81 and the B-connection channel 82 of the wake control valve, for which the outside of the valve housing 52 are inserted into these, are pressure-tight on these outside sides plugged.

The configuration of the overrun control valve 39 shown in FIGS. 3b and 3c corresponds, regardless of the arbitrarily chosen orientation of the longitudinal center planes 93 and 94 of the P control grooves 88 and 89 and of the T control grooves 91 and 92, to that Shown schematically in FIG. 1b, blocking position 0 of the follow-up control valve 39, in which there is no overlap - positive overlap - of the A control channels 96 and 97 and the B-control channels 102 and 103 of the valve bushing 54 with the P-control grooves 88 and 89 and the T-control grooves 91 and 92 of the central valve piston 57 are given and thus - at least on average over time - the A-connection channel 81 and the B connection channel 82 of the housing 52 of the overflow control valve 39 are blocked off against the P control grooves 88 and 89 and the T control grooves 91 and 92 of the central valve piston 57.

This configuration of the overrun control valve corresponds to the equality of the setpoint and actual value of the azimuthal position of the rotor 14 of the gerotor motor 11. It is - at the same time inevitable - achieved at the end of a positioning process and therefore also forms the initial position for a subsequent positioning process, which always begins from the position of the follow-up control valve 39 shown in FIGS. 3b and 3c.

Starting from the blocking position 0 of the overrun control valve 39 shown in FIGS. 3b and 3c, the valve bushing 54 is actuated by actuating the stepping motor 41 in the direction of the arrow 112 in FIG. 3b, ie according to the illustration of this figure, in FIG Turned clockwise relative to the central valve piston 57, the A-control channels 96 and 97 of the valve bush 54 come into positive overlap of their flow cross-sections with the P-control grooves 88 and 89 of the central valve piston 57 (FIG. 3d), while the B- Control channels 102 and 103 of the valve bush 54 of the follow-up control valve 39 in positive overlap of their control cross sections with the T-control grooves 91 and 92 of the central valve piston 57 of the overrun control valve 39 (FIG. 3e), so that ε assumes a functional position I shown schematically in FIG. 1, in which the gerotor motor 11, depending on the design of the feedback gear 58 or 58 'is driven in the opposite direction to the direction of rotation of the valve bush 54 or with the direction of rotation which is of the same sense to the direction of rotation thereof.

On the other hand, starting from the initial position of the valve bushing 54 shown in FIGS. 3b and 3c, this is driven by the control of the stepping motor 41 in the opposite direction of rotation, ie in the direction of rotation of the arrow 113 of FIGS. 3f and 3g, so that A-control channels 96 and 97 of valve sleeve 54 in positive overlap of their flow cross sections with T-control grooves 91 and 92 of central valve piston 57 (FIG. 3f) and B-control channels 102 and 103 of valve sleeve 54 in positive overlap with the cross sections the P control grooves 88 and 89 of the central valve piston 57 of the follow-up control valve 39 (FIG. 3g), which thereby reaches its functional position II shown schematically in FIG. 1 a, in which the rotor 14 of the gerotor 11 in the direction of rotation , which the follow-up control valve 39 conveys in its functional position I, is driven in the opposite direction of rotation, again depending on how that Feedback gear 58 or 58 'of the drive unit 10 is designed. In order to achieve the backlash of the feedback gear 58, which is suitable for a precise function of the drive unit 10, a tensioning device which mediates the function of a torsion spring, generally designated 114, is provided, which rotatably rests on the central valve piston 57 on the valve sleeve 54 is connected to the output shaft of the stepping motor 41, exerts azimuthally supported torque, due to which the internal toothing 67 of the ring gear connected non-rotatably to the central valve piston 57 is reliably held in one-sided system with the teeth of the pinion 64 which are in engagement with it. This is connected to the rotor 14 of the gerotor motor.

This tensioning device 114 comprises a helical spring 116 under tension, which on an azimuthal area spanning approximately 300 ° from an outer, concave groove 117 has an end portion 118 which is only slightly extended in the axial direction and protrudes from the central housing bore 53 on the transmission side the valve bush 54 is received. The radius of curvature of the groove is slightly larger than that of the spring coils, which are received with a radially inner 180 ° area by this concave groove 117 and are supported at the bottom thereof. The short end section 118 of the valve bush 54 extends through a bore step 119, the diameter of which is slightly larger than the outer diameter, in relation to the central bore 53, in which the valve bush is arranged so that it can be rotated in a pressure-tight manner over sections of its length the helical spring 116, the radial clear width between the bore step 119 and the outer lateral surface of the end section 118 of the valve bushing 54 bearing the helical spring 116 remaining annular gap 121 is smaller than the diameter of the individual spring coils, which is around 0 with a spring wire thickness .2 mm is approximately 2 mm. As a result, the coil spring 116 is adequately secured against axial displacement out of the annular gap 121.

In the central valve piston 57, a stop pin 122 is fixedly inserted in the area 57 'from the end section 118 of the valve bushing 54 in the azimuthal area of 300 ° coaxially, emerging from the central housing bore 53 on the transmission side, which on one side radially into the "Free" annular gap region 121 'protrudes, the azimuthal width 0 of which is determined by the azimuthal distance of the radial end faces 123 and 124, which extend in the axial direction over the depth of the end section 118 of the valve bush 54 carrying the coil spring 116.

This stop pin 122 is oriented so that the radial plane containing its central longitudinal axis 126 and the central longitudinal axis 19 of the follow-up control valve 39 halves the angle that the radial plane which the common central longitudinal axis 106 of the A control channels 96 and 97 contains and enclose the radial plane, which contains the common central longitudinal axis 104 of the B control channels 102 and 103. With this orientation of the central longitudinal Axis 126 of the stop pin 122, in the - blocking - central position of the follow-up control valve 39, has the same azimuthal distance from the two radial end faces 123 and 124 of the end section 118 of the valve bushing 54 which supports the helical spring 116, so that, starting from this middle position, for the control of the wake control valve in its alternative flow positions I and II are available in terms of the same control angle.

One end 127 of the helical spring 116 is attached to the free end section 122 'of the stop pin 122, while the other end 128 is fixed to the region of the 300 ° sector-shaped jacket section 118 of the valve bush 54 which is approximately 300 ° away in the direction of travel of the spring.

With regard to its dimensioning and its preload on the holding torque of the stepping motor 41 in its de-energized state, and the holding torque of the gerotor motor 11 when the pressure supply is switched off, the helical spring 116 is matched in such a way that there is between the valve bush 54 and the central valve piston 57 - as a result, between the rotor of the stepper motor 41 and the rotor 14 of the gerotor motor - due to the spring preload, the permanently effective torque is far from sufficient to twist it against one another in a standstill phase of the drive unit. After the gerotor motor 11 has stopped and it has reached its blocking central position, the overrun control valve 39 cannot, as it were, operate in reach. The bias of the helical spring 116 of the tensioning device 114, which "tends" to push the overrun control valve 39 into one of these functional positions I or II, is only sufficient for the fact that the position feedback of the rotor 14 of the gerotor Motorε 11 required freedom of play of the respective feedback gear 58 or 58 'is guaranteed.

Claims

claims
1. Electrohydraulic control device for a rotary hydraulic motor, in which the rotor carries out two superimposed rotary movements, one of which takes place eccentrically to the central longitudinal axis of the stator, and the other of which rotates around the rotor-fixed central one running parallel to the longitudinal axis of the stator Longitudinal rotor axis, which thereby executes a circular movement around the central longitudinal axis of the stator along a circular path, and the rotary movements of the rotor are converted into a uniaxial rotary movement of the output shaft of the motor by means of a cardan shaft coupling the output shaft to the rotor , as well as a control circuit working with a variable position setpoint specification, in which a uniaxially rotatably driven, mechanically driven mechanical sensor element is provided to measure the position value et that the sensor element (57) used for position value measurement with the rotor by means of its own, in addition to the drive train formed by the rotor (14), the cardan shaft (34) and the output shaft (36), Over which a predominant part of the engine torque is directed as the useful torque, the provided sensor drive line (58; 58 ') is motion-coupled.
2. Control device according to claim 1, thereby characterizes that the gear (58) which conveys the conversion of the rotor movements into the uni-axial rotary movements of the mechanical sensor element (57) of the position-actual value detection device which can be monitored is designed as a trochoid gear.
Control device according to claim 3, characterized in that the conversion of the rotary movements of the rotor (14) of the rotary hydromotor (11) into the monitorable uniaxial rotary movements of the mechanical sensor element of the position actual value detection device Mediating additional gearboxes are constructed as hypotrochoid gearboxes (58 '), the ring gear (63') of which is connected to the rotor (14) of the hydraulic motor (11) in a rotationally fixed manner.
Control device according to claim 2, characterized in that the conversion of the rotary movements of the rotor (14) of the rotary hydraulic motor (11) into the monitorable uniaxial rotary movements of the mechanical sensor element (57) of the position-sensing value-detection device transmits gears as hypotrochoid devices. Gearbox is formed, the ring gear (63) of which is rotatably connected to the uniaxially rotating sensor element (57) of the position / value detection device.
Control device according to one of claims 1 to 4, characterized in that the daε further gears (58; 58 ') uniaxially rotatable element (57) of the feedback device functional element of the mechanical actual value feedback device of a follow-up control valve (39) provided for motion and position control of the rotary hydraulic motor (11) , that works with an electrically controllable setpoint specification by means of a step motor or an AC motor (41).
6. Control device according to claim 5, characterized ge indicates that the follow-up control valve (39) is designed as a rotary slide valve with rotatably driven piston and housing elements (54, 57), the azimuthal deflection of which is impact-limited against one another.
7. Control device according to claim 6 in combination with claim 3, characterized in that the centrally arranged valve piston (57) of the post-flow control valve (39) with a pinion (64 ') is provided with the ring gear (63 ') of the rotor (14) is in meshing engagement, and the housing (54) of the overrun control valve (39) can be driven by means of the setpoint specification electric motor.
8. Control device according to claim 6 in combination with claim 5, characterized in that the centrally arranged valve piston (57) can be driven by means of the setpoint specification motor (41) and the valve housing (54) with the pinion (64 ) of the rotor (14) of the Hydromotorε (11) meshing Ring gear (63) is provided.
9. Control device according to claim 7 or claim 8, characterized in that the housing of the follow-up control valve (39) is designed as a valve bushing (54), which in turn is located in a central through bore (53) of one with the motor housing (28). firmly connected valve housing block (52) of the overrun control valve (39) with a sealing sliding seat is rotatably arranged.
10. Control device according to one of claims 6 to 9, characterized in that a bracing device (114) is provided which interposes a valve bushing (54) and the piston (57) of the follow-up control valve (39) permanently effective torque ment that is smaller in magnitude than the holding torque of the setpoint specification motor (41) in its de-energized state and also smaller than the holding torque of the rotary hydraulic motor when the pressure supply is switched off.
PCT/EP1996/002818 1995-06-27 1996-06-27 Electrohydraulic control arrangement for a rotary hydraulic motor WO1997001711A1 (en)

Priority Applications (2)

Application Number Priority Date Filing Date Title
DE19522768.9 1995-06-27
DE1995122768 DE19522768A1 (en) 1995-06-27 1995-06-27 Electro-hydraulic control device for a rotary hydraulic motor

Applications Claiming Priority (3)

Application Number Priority Date Filing Date Title
EP19960923960 EP0835382B1 (en) 1995-06-27 1996-06-27 Rotary hydraulic motor with electrohydraulic control arrangement
DK96923960T DK0835382T3 (en) 1995-06-27 1996-06-27 Hydraulic rotary motor with electro-hydraulic control device
DE1996501508 DE59601508D1 (en) 1995-06-27 1996-06-27 Rotational hydromotor with electrohydraulic control device

Publications (1)

Publication Number Publication Date
WO1997001711A1 true WO1997001711A1 (en) 1997-01-16

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PCT/EP1996/002818 WO1997001711A1 (en) 1995-06-27 1996-06-27 Electrohydraulic control arrangement for a rotary hydraulic motor

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EP (1) EP0835382B1 (en)
AT (1) AT178123T (en)
DE (1) DE19522768A1 (en)
DK (1) DK0835382T3 (en)
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* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE19945122B4 (en) * 1999-09-21 2004-08-12 Sauer-Danfoss Holding Aps Hydraulic control device
US6439101B1 (en) * 1999-10-13 2002-08-27 Teijin Seiki Co., Ltd. Electro-hydraulic servomotor
CN104847257B (en) * 2015-04-20 2017-12-08 江汉石油钻头股份有限公司 A kind of screw drilling tool motor
CN104847258B (en) * 2015-04-20 2017-12-08 江汉石油钻头股份有限公司 A kind of all-metal screw drilling tool

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DE2110863B1 (en) * 1971-03-08 1972-08-31 Danfoss As Parallel and internal-axis rotary piston engine
US4494915A (en) * 1979-06-25 1985-01-22 White Hollis Newcomb Jun Hydrostatic steering unit with cylindrical slide member within clindrical valve sleeve

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DE4015101A1 (en) * 1990-05-11 1991-11-14 Eckehart Schulze Hydraulic drive device
DE9308025U1 (en) * 1993-05-27 1993-07-29 Moog Gmbh, 7030 Boeblingen, De

Patent Citations (2)

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Publication number Priority date Publication date Assignee Title
DE2110863B1 (en) * 1971-03-08 1972-08-31 Danfoss As Parallel and internal-axis rotary piston engine
US4494915A (en) * 1979-06-25 1985-01-22 White Hollis Newcomb Jun Hydrostatic steering unit with cylindrical slide member within clindrical valve sleeve

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EP0835382A1 (en) 1998-04-15
AT178123T (en) 1999-04-15
EP0835382B1 (en) 1999-03-24
DE19522768A1 (en) 1997-01-02
DK835382T3 (en)
DK0835382T3 (en) 2000-06-05

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