Electrohydraulic actuator

Abstract

The invention relates to an electrohydraulic actuator comprising a hydraulic piston, a spindle drive for moving the piston along an axis of rotation, wherein the spindle drive comprises a spindle and a nut, and a worm wheel, which is coaxially attached to the nut. Furthermore, an electric motor having a worm, which meshes with the worm wheel, is provided. The nut is radially and axially supported by means of a bearing. In addition, there is a single-armed rotation-prevention mechanism in order to prevent the spindle from rotating. An angle with respect to the axis of rotation is included between a point of action of the worm on the worm wheel and the effective direction of the rotation-prevention mechanism, wherein the angle is selected in such a way that, for a predetermined torque acting on the nut, a radial force acting on the piston is minimized.

Claims

1. An electrohydraulic actuator (100) comprising a hydraulic piston (115); a spindle drive (140), comprising a spindle (145) and a nut (150), for moving the piston (115) along an axis of rotation (175); a worm wheel (170), which is attached coaxially to the nut (150); an electric motor (130) having a worm (165), which meshes with the worm wheel (170); a radial bearing (155) for supporting the nut (150), and a single-armed rotation-prevention mechanism (185) for securing the spindle (145) against rotation, wherein an angle () with respect to the axis of rotation (175) is included between a point of action (205) of the worm (165) on the worm wheel (170) and a direction of action (A) of the rotation-prevention mechanism (185), and wherein the angle () minimizes, for a predetermined torque acting on the nut (150), a radial force (F.sub.piston) acting on the piston (115).

2. The actuator (100) as claimed in claim 1, wherein the radial bearing (155) allows a predetermined tilt of the nut (150) relative to the axis of rotation (175).

3. The actuator (100) as claimed in claim 1, wherein the radial bearing (155) is arranged between the nut (150) and the rotation-prevention mechanism (185) in an axial direction.

4. The actuator (100) as claimed in claim 1, wherein the angle () is determined with respect to a predetermined axial position of the rotation-prevention mechanism (185) along the spindle (145).

5. The actuator (100) as claimed in claim 1, wherein the angle () is determined with respect to a predetermined friction coefficient () between the worm (165) and the worm wheel (170).

6. The actuator (100) as claimed in claim 1, wherein the torque acting on the nut (150) is determined by way of an axial force acting on the piston (115) at a predetermined operating point.

7. The actuator (100) as claimed in claim 2, wherein the radial bearing (155) is arranged between the nut (150) and the rotation-prevention mechanism (185) in an axial direction.

8. The actuator (100) as claimed in claim 7, wherein the angle () is determined with respect to a predetermined axial position of the rotation-prevention mechanism (185) along the spindle (145).

9. The actuator (100) as claimed in claim 8, wherein the angle () is determined with respect to a predetermined friction coefficient () between the worm (165) and the worm wheel (170).

10. The actuator (100) as claimed in claim 9, wherein the torque acting on the nut (150) is determined by way of an axial force acting on the piston (115) at a predetermined operating point.

Description

BRIEF DESCRIPTION OF THE DRAWINGS

(1) The invention is now described in greater detail with reference to the attached figures, of which:

(2) FIG. 1 shows an electrohydraulic actuator;

(3) FIG. 2 the actuator from FIG. 1 in another view;

(4) FIG. 2A a detail view of a portion of the actuator of FIG. 2;

(5) FIG. 3 shows a force and torque diagram for the actuator in FIGS. 1 and 2, and

(6) FIG. 4 shows another force and torque diagram for the actuator in FIGS. 1 and 2.

DETAILED DESCRIPTION

(7) FIG. 1 shows an electrohydraulic actuator 100 for supplying a hydraulic pressure, in particular for actuating a brake or a clutch on board a motor vehicle. The actuator 100 comprises a hydraulic module 105, which comprises a cylinder 110 and a piston 115 accommodated therein. To seal the piston 115 relative to the cylinder 110, a piston ring 120 or some other radial seal is preferably provided. The cylinder 115 contains a hydraulic fluid 125, which is put under pressure when the piston 115 is pushed into the cylinder 110. The actuator 100 furthermore comprises an electric motor 130 and a gear unit 135 for transmitting a movement of the electric motor 130 to the hydraulic module 105. The gear unit 135 comprises a spindle drive 140, which, for its part, comprises a spindle 145 and a nut 150, a radial bearing 155 and a worm gear 160, which comprises a worm 165 and a worm wheel 170. The worm 165 meshes in the worm wheel 170, wherein an axis of rotation of the electric motor 130 includes an angle with the plane of representation, with the result that a point of action of the worm 165 on the worm wheel 170 is not visible in FIG. 1. The worm wheel 170 is attached positively and coaxially to the nut 150 and both elements can be rotated about an axis of rotation 175. The axis of rotation 175 is oblique with respect to the axis of rotation of the electric motor 130. In the radial direction, the radial bearing 155 is preferably arranged between the nut 150 or the point of action of the worm 165 on the worm wheel 170 and the rotation-prevention mechanism 185.

(8) The nut 150 with the worm wheel 170 is supported by means of the radial bearing 155, e.g. relative to a housing 180. The radial bearing 155 is preferably embodied as a rolling bearing, in particular as a ball bearing. Here, the radial bearing 155 preferably allows a certain tilting of the nut 150 relative to the direction of motion of the piston 115. The axis of rotation 175 can thus assume a predetermined maximum tilt angle relative to the direction of motion of the piston 115. By allowing the tilt angle, it is possible to compensate an offset misalignment or an angular misalignment between the axis of rotation 175 and the direction of motion of the piston 115.

(9) The spindle 145 is attached to the piston 115 at a first end. In the embodiment shown, the spindle 145 carries a ball at the first end, and the cylinder 115 has a ball socket corresponding thereto, allowing axial forces to be exchanged between the spindle 145 and the piston 115, while tilting of the piston 115 by the spindle 145 is prevented. A rotation-prevention mechanism 185 is attached at its opposite, second end of the spindle 145. The rotation-prevention mechanism 185 comprises a lever arm 190, which extends in a radial direction, and an axial groove 195, in which a radially outer end of the lever arm 190 is received. Other embodiments which prevent rotation of the lever arm 190 with respect to the housing 180 and simultaneously allow an axial motion are likewise possible, e.g. a slide rail. To assist the tilting of the nut 150, it is preferred that the rotation-prevention mechanism 185 should not absorb any radial or tilting forces.

(10) A coordinate system in FIG. 1 is shown only in part. The y direction extends parallel to the axis of rotation 175 and the z direction extends parallel to the shaft of the electric motor 130. The x direction forms a right-handed coordinate system with the y direction and the z direction.

(11) The following dimensions are furthermore defined in FIG. 1: a axial distance between the point of action of the spindle 145 on the piston 115 and the point of action of the worm 165 on the worm wheel 170; b axial distance between the radial bearing 155 and the point of action of the worm 165 on the worm wheel 170; c axial distance between the radial bearing 155 and the rotation-prevention mechanism 185; d radial distance between the axis of rotation 175 and the point of action between the worm 165 and the worm wheel 170 (not visible), and e radial distance between the central point of support of the rotation-prevention mechanism 185 and the axis of rotation 175.

(12) FIG. 2 shows the actuator 100 in another view. The axis of rotation 175 or axis of motion of the piston 115 in the cylinder 110 extends perpendicularly to the plane of representation. In the plane of representation, the axis of rotation of the electric motor 100, which extends obliquely to the axis of rotation 175, extends in a horizontal direction. The worm 165 and the worm wheel 170, which mesh with one another, are shown in a partial section. Here, a point of action 205 refers to the point at which a tensile or shear force is transmitted between the worm 165 and the worm wheel 170. The point of action 205 is at a predetermined distance and has a predetermined alignment with respect to the axis of rotation 175.

(13) In a similar way, the rotation-prevention mechanism 185 is partially cut away. The bearing point of the lever arm 190 on the groove 195 has a predetermined rotational alignment with respect to the axis of rotation 175 and is at a predetermined distance therefrom. The rotational alignment is also referred to as the direction of action A. An angle between the direction of action A and a radius connecting the axis of rotation 175 to the point of action 205 is referred to as . A second angle complements to give 180. The direction of action A is shown relative to a center line of the lever arm 190. For more precise determination, the direction of action A can instead pass through the effective contact area between the lever arm 190 and the groove 195 and the axis of rotation 175. In the case of the usual loading of the actuator 100, the effective contact area is that shown on the right in FIG. 2.

(14) FIG. 3 shows a force and torque diagram for the actuator 100 in FIGS. 1 and 2. The alignment corresponds to that in FIG. 1. From the top down, the forces shown relate to the piston 115, the worm wheel 170, the radial bearing 155 and the rotation-prevention mechanism 185. The forces of the rotation-prevention mechanism 185, which are shown in an ellipse in dashed lines, have been rotated by the angle out of the plane of the drawing. For further explanation, the following designations are used:

(15) F.sub.x.sub._.sub.piston 305

(16) F.sub.y.sub._.sub.piston 310

(17) F.sub.z.sub._.sub.piston 315

(18) F.sub.x.sub._.sub.worm wheel 320

(19) F.sub.y.sub._.sub.worm wheel 325

(20) F.sub.z.sub._.sub.worm wheel 330

(21) F.sub.x.sub._.sub.bearing 335

(22) F.sub.y.sub._.sub.bearing 340

(23) F.sub.z.sub._.sub.bearing 345

(24) F.sub.x.sub._.sub.rot.-prev.mech. 350

(25) F.sub.y.sub._.sub.rot.-prev.mech. 355

(26) F.sub.z.sub._.sub.rot.-prev.mech. 360

(27) F.sub.rot.-prev.mech. 365

(28) FIG. 4 shows another force and torque diagram 400 for the actuator 100 in FIGS. 1 and 2. The illustration corresponds to that in FIG. 3 from a different perspective. The selected view corresponds to that in FIG. 2.

(29) In order to minimize the forces acting radially on the piston 115, F.sub.piston must be minimized. Here:
F.sub.piston={square root over (F.sub.x.sub._.sub.piston.sup.2+F.sub.z.sub._.sub.piston.sup.2)}
F.sub.x.sub._.sub.rot.-prev.mech.=F.sub.rot.-prev.mech.*sin
F.sub.z.sub._.sub.rot.-prev.mech.=F.sub.rot.-prev.mech.*cos
F.sub.y.sub._.sub.rot.-prev.mech.=F.sub.rot.-prev.mech.*

(30) =sliding friction coefficient

(31) Force and torque balance:
F.sub.x=F.sub.x.sub._.sub.pistonF.sub.x.sub._.sub.worm wheelF.sub.x.sub._.sub.bearing+F.sub.x.sub._.sub.rot.-prev.mech.=0
F.sub.y=F.sub.y.sub._.sub.pistonF.sub.y.sub._.sub.worm wheel+F.sub.y.sub._.sub.bearingF.sub.y.sub._.sub.prev.mech.=0
F.sub.z=F.sub.z.sub._.sub.piston+F.sub.z.sub._.sub.worm wheel+F.sub.z.sub._.sub.bearing+F.sub.z.sub._.sub.rot.-prev.mech.=0
M.sub.x=F.sub.z.sub._.sub.piston(a+b)+F.sub.z.sub._.sub.worm wheelbF.sub.z.sub._.sub.rot.-prev.mech.c+F.sub.y.sub._.sub.rot.-prev.mech.e sin =0
M.sub.y=F.sub.z.sub._.sub.worm wheeld+F.sub.rot.prev.meche=0
M.sub.z=F.sub.x.sub._.sub.piston(a+b)+F.sub.x.sub._.sub.worm wheelbF.sub.y.sub._.sub.worm wheeld+F.sub.x.sub._.sub.rot.-prev.mech.+F.sub.y.sub._.sub.rot.-prev.mech.e cos =0

(32) Given, known variables: The maximum effective piston force F.sub.y.sub._.sub.piston The maximum effective gearing forces F.sub.x.sub._.sub.worm wheel, F.sub.y.sub._.sub.worm wheel, F.sub.z.sub._.sub.worm wheel The installation space conditions a, b, c, d, e The sliding friction coefficient , assumed to be 0.15 for example.

(33) Unknown variables:

(34) F.sub.x.sub._.sub.piston

(35) F.sub.z.sub._.sub.piston

(36) F.sub.rot.-prev.mech.

(37)

(38) F.sub.x.sub._.sub.bearing

(39) F.sub.y.sub._.sub.bearing

(40) F.sub.z.sub._.sub.bearing

(41) The forces F.sub.x.sub._.sub.piston and F.sub.z.sub._.sub.piston are specified as a magnitude limitation. In particular, a resultant force acting radially on the piston 115 and composed of F.sub.x.sub._.sub.piston and F.sub.z.sub._.sub.piston, F.sub.piston can be limited in magnitude as a design stipulation. This design stipulation can be obtained from empirical values in respect of the wear behavior of the piston 115 and of the cylinder 110. For example, it can be stipulated that the radial piston force F.sub.piston must not exceed a certain amount when the actuator 100 is at a predetermined operating point. The operating point can be given by a hydraulic operating pressure of the hydraulic fluid 125 or a position of the spindle 145 along the axis of rotation 175, for example.

(42) In order to minimize the radial piston forces F.sub.piston, needs to be optimized. For example, can be set to different values in a test series, while the system of six equations indicated above is resolved in respect of the remaining six unknown variables. The system of equations is therefore determined and there is always a unique solution. is then chosen in such a way that F.sub.x.sub._.sub.piston and F.sub.z.sub._.sub.piston are minimized, with the result that the force acting radially on the piston 115, F.sub.piston is likewise minimized.