WO2019214835A1

WO2019214835A1 - Piston-cylinder system having separate bearing and sealing region

Info

Publication number: WO2019214835A1
Application number: PCT/EP2018/078899
Authority: WO
Inventors: Heinz Leiber
Original assignee: Ipgate Ag
Priority date: 2018-05-09
Filing date: 2018-10-22
Publication date: 2019-11-14

Abstract

The invention relates to a piston-cylinder system, in particular for an actuator unit or brake system, comprising a cylinder (7) having a cylinder bore having an internal diameter (d4) and a piston (5) axially mounted therein so as to be movable, at least one seal (D1, D2) being arranged in a first axial piston region (5_B1) between the outer wall of the piston (5) and the inner wall of the cylinder, characterised in that the piston has a second piston region (5_B2), the external diameter (d2) of which is greater than the external diameter (d3) of the first axial piston region (5_B1).

Description

Piston-cylinder system with separate bearing and sealing area

The present invention relates to a piston-cylinder system, in particular for a control unit or brake system, with a cylinder with a cylinder bore having an inner diameter and a piston axially displaceably mounted therein, wherein in a first axial piston area between the Außenwan- the piston and the inner wall of the cylinder at least one seal is arranged.

State of the art

Motor drives with spindle for adjusting an adjusting element, in particular for adjusting an actuator or a piston of a hydraulic piston drive are widely used. The rotational movement of the rotor is thereby transferred to a spindle or spindle nut, wherein the spindle or spindle nut is connected to the adjusting element, in particular in the form of an actuator or piston, and moves it linearly back and forth.

For high efficiencies, a ball screw KGT is often used. The torque M _d generated by the motor must be supported so that the adjusting element does not rotate with it. For this purpose, an anti-rotation device is used, which acts either directly on the adjusting element or on the spindle nut, provided that the spindle is non-rotatably connected to the rotor. The anti-rotation device can also act on the spindle if the spindle nut is connected in a rotationally fixed manner to the rotor. When using a helical gear, preferably a ball screw, the following listed problems exist:

A) The drive generates over the various tolerances an eccentricity of the motor axis to the axis of the rotatable spindle or spindle nut, which on the adjusting element or the piston, the helical gear, consisting of spindle and spindle nut, and possibly seals and Ver - is transmitted rotation lock.

B) When the eccentricity is transmitted to the adjusting element or the piston, depending on the elasticity of the transmission, a transverse force is produced on the adjusting element or the piston.

C) The lateral force described in B) generates friction and wear on the adjusting element or the piston.

D) If the piston wears, z. B. form scratches, the sealing effect is no longer secure, which collides in particular with fail-operation requirements FO.

E) On the anti-rotation also affects the eccentricity and in addition to be supported motor torque M _d . The applied forces and the associated friction usually produce a high degree of wear. The aforementioned forces can also generate additional lateral forces on the adjusting element or the piston.

F) The wear by E) can also get on the piston barrel, so that the seal moves and this can fail, which is also critical in fail-operation requirements.

G) Dependence of the functions of temperature and tolerances.

H) Complexity, number of parts for spindle drive with motor and anti-rotation protection and tolerance chains. In addition to the tolerances, which determine the eccentricity of the rotating parts, the deviation from the motor, spindle and piston axis also has an effect on the spindle drive. As an example, the coupling of the engine to the piston housing can be mentioned here. Typically, the engine has a flange that dips into the piston housing, with tolerances for flange and piston housing. These tolerances result in a minimum and maximum clearance, which results in a corresponding offset of the axes to each other. To other components of the spindle drive results in a considerable offset, which can be up to a factor of five more than the eccentricity. This offset can also affect the above-mentioned problems A) to F) with a corresponding construction.

Spindle drives are described numerous. Thus, the documents listed below all show electromotively driven pistons, wherein a helical gear in the form of a spindle drive is interposed between the electric drive and the piston. In order to save a lot of unnecessary descriptions and explanations, the documents will each refer to the problems that occur in the following.

From DE 10 2008 059 862 Al an electrohydraulic brake system for motor vehicles is known in which the rotor is firmly connected to the spindle nut, and the spindle with the piston are integrally formed. The rotor is mounted on the housing together with the spindle nut by means of two bearings, the rotation lock being arranged in the end of the spindle-piston component facing away from the piston. In this construction, in particular, the problems A), B), C), D), E) and F) arise.

DE 11 2009 004 636 likewise describes a brake system in which a piston of a piston-cylinder system is connected to the driven spindle via a tappet. Here, in particular, the problems A), B), C), D), E), F) and G) occur, whereby altogether lower loads occur, since the pistons are connected to the spindle via the tappet, so that z. B. the eccentricity and shear forces are reduced by the formed by the plunger joint piece. From DE 10 2011 106 626 Al a highly dynamic construction engineering drive for a piston-cylinder unit is known, in which also the problems A), B), C),

D), E), F) and G) play a role, whereby the loads occurring at the anti-rotation device are also reduced by an elastic sleeve and a roller.

In the electric drive described in DE 10 2013 221 158 A1, the problems A) to G) occur, whereby due to the eccentricity of the rotating parts high loads and forces occur due to a low elasticity and the rotation lock.

The DE 10 2014 212 409 A1 known pressure generator for a hydraulic vehicle brake has a dual-mounted rotor, which drives a spindle nut via a planetary gear. Spindle and piston are integrally formed, wherein the bearing of the piston takes place solely by its lateral surface and the rotation is effected by a rod engaging in the piston end, in cross-section hexagonal, rod. Even with this pressure generator problems A) to G) occur. The same applies to the known from DE 10 2015 222 286 Al hydraulic unit. Here occur high lateral force loads on the piston, which has an increased wear also for the cylinder inner wall and the seals result. The problems A) to G) can also occur with DE 10 2017 211 587 A1 and with the brake pressure control unit known from DE 10 2016 208 367 A1. In DE 10 2016 208 367 A1, an elastically deformable preload element is used to compensate for larger tolerances.

In piston-cylinder systems, there is also the problem that the seals are exposed to excessive forces.

Object of the invention

The object of the invention is to provide a piston-cylinder system in which the seals are not exposed to great forces and the piston is sufficiently well stored in the cylinder bore. Another task of The present invention is to reduce or avoid tumbling of the rotor at stroke start and / or stroke direction change of the piston.

Solution of the task

This object is achieved with a piston-cylinder system with the features of claim 1. Advantageous embodiments of the piston-cylinder system according to claim 1 result from the features of subclaims 2 to 13. By separating the sealing region and the bearing region, the play between piston and cylinder inner wall being greater in the sealing region than in the bearing region, It can advantageously be prevented that the piston does not come into contact with the cylinder inner wall in this area and possible damage to the seals, in particular in the form of sealing rings, is prevented. In this case, the entire lateral surface of the second piston area acts as a bearing surface and is interrupted - if at all - only in the circumferential direction by the at least one projection of the anti-twist device.

The disadvantageous tumbling of the piston can also advantageously be prevented by exerting a force acting in the axial direction on the piston and / or the rotor by means of a spring element. It is also possible to reduce the tumbling by means of a radial support of the rotor or to eliminate completely.

The piston-cylinder system according to the invention can be used particularly advantageously in conjunction with an electric motor-driven helical gear, which in turn adjusts the piston. In this case, it is particularly advantageous if the bearings supporting the rotor, the helical gear and the piston are arranged as far away from each other as possible.

In this case, there is advantageously no further bearing between the first bearing and the end region or a radial sliding bearing arranged in this end region for radially supporting the rotor and the helical gear. By this training already results in a relatively large transverse elasticity, through which the wear is significantly reduced by eccentricity of the rotating parts. By providing a transverse elasticity in the rotor, helical gear and / or the adjusting element as non-deformable elasticity, the wear can advantageously be further reduced, since the transverse forces are advantageously reduced by the elasticity. Thus, it is advantageous if the rotor is formed at least in a region between the first bearing and the input of the helical gear transverse elastic to the axis of rotation, in particular a spring element or a region with higher elasticity. As a result, virtually all the above-described problems A) to G) are solved or greatly reduced. Also, the helical gear according to the invention with drive and adjusting element has a low complexity. The solution can be used both for rotating spindle or spindle nuts, with rotating spindle nuts usually requiring a little more length but advantageously no grease being thrown away from the revolving spindle and additional measures such as additional guard ring on the SM being required. Advantageously, a ball screw KGT is used because of the good efficiency.

Thus, various spindle gearboxes, in particular ball screw gearboxes, can also be used. It is also possible that the adjusting element is an actuator or a coupling part to a drive, which is moved back and forth on a web and even something drives or adjusted. Preferably, however, a piston is moved back and forth via the helical gear, wherein the piston is displaced in a cylinder for pressure generation and pressure reduction or for holding pressure in one or more hydraulic travel. The piston can be designed as a single stroke piston, which limits only one pressure chamber, or as a double-stroke piston, which separates two pressure chambers from each other in a sealing manner. Of course, other piston-cylinder systems can likewise be driven via the helical gearbox driven by the invention.

The first bearing can advantageously be arranged between the rotor winding and / or the permanent magnet bearing part of the rotor and the input of the helical gear, being seen as the input of the helical gear in the context of the invention, the rotor driven by the component of the helical gear, which the Spindle nut or the spindle can be. It is also possible that a transformer is mounted by means of the first bearing and is rotatably connected to the input of the helical gear, wherein the transformer is driven by a rotor of an engine and a gearbox between geschalte¬.

It is especially advantageous if, additionally in the region of the anti-rotation lock, a radial sliding bearing is arranged or the anti-twist device itself is or is additionally a radial plain bearing. As a result, both the radial support and the rotation is ensured in a small space.

At the beginning of a piston stroke or upon reversal of the stroke, no counterforce or no axial centering force acts over a short time or for a short stroke movement on the first bearing and / or on the ball screw gear KGT, which leads to a tumbling of the rotor can. The invention provides for the reduction of wobbling two possible alternative solutions, but which can also be provided together. Thus, an additional radial support, in particular with a small clearance for the rotor and / or a piston return spring can be provided.

Description of the figures

Hereinafter, various possible embodiments of the invention will be explained in more detail with reference to drawings.

Show it :

Fig. 1: shows the spindle drive with rotating spindle with motor and

Rotor bearing (bearing LI), elastic rotor and second bearing on the anti-twist device and piston seals;

Fig. La: shows an anti-rotation with sliding guide;

2 shows an embodiment with a radial support element and / or a piston return spring for preventing a tumbling motion of the rotor; 3 shows another possible embodiment with a spring element for exerting a force acting in the axial direction on the rotor and / or its bearing;

Fig. 4: further possible embodiment, one of the in the figures 1 to

3 illustrated and described measure in the form of the Fe derelementes (Figure 3), the return spring (Figure 2) and / or the radial bearing (Figure 2) can be provided;

Fig. 4a: shows an anti-rotation with sliding guide;

Fig. 4b: shows an anti-rotation with roller guide; 5 shows a spindle drive with revolving spindle nut, in which optionally one of the measures shown in Figures 1 to 3 and described in the form of the spring element (Figure 3), the return spring (Figure 2) and / or the radial bearing (Fi gur 2) can be provided; Fig. 6 shows a Doppelhubkolben (DHK) with a slide bearing and

Rotation within the piston, in which optionally one of the measures shown in Figures 1 to 3 and described in the form of the spring element (Figure 3), the return spring (Figure 2) and / or the radial bearing (Figure 2) are provided that can;

7 shows a DHK with second bearing on the piston and an anti-rotation lock with Oldham coupling, in which optionally one of the measures illustrated and described in FIGS. 1 to 3 in the form of the spring element (FIG. 3), the return spring (Figure 2) and / or the radial bearing (Figure 2) can be provided; ;

Fig. 7a: shows anti-rotation, which is realized by an Oldham coupling. Figure 1 shows a schematic representation of the essential components of the motor spindle drive with motor housing 1, stator with winding 2, rotor 3, spindle S, spindle nut SM, helical gear 4, which as a ball screw drive (KGT) is formed, piston 5 and anti-rotation VS , The rotor 3 is mounted in the first ball bearing LI and is connected to a nut 10 with the spindle S. As the state of the art shows, the kinematics of rotation and translation into translation are very complex. Among other things, the eccentricity and its effect on the anti-rotation protection must be taken into account with the total balance of forces of the radial and axial forces with resulting frictional forces. The rotor 3 is fixed by means of the screw 10 via a snug fit on the spindle cap Sz. In addition to the frictional connection via the frictional torque of the axial force of the nut 10, it is also possible to use an additional frictional connection, for example by means of a pin or ball (not shown) between the spindle S and the rotor flange 3f. The fit between pin and rotor 3 determines the first tolerance. The fit between rotor 3 and ball bearing LI the second tolerance. Both fits determine next to the impact of fit to the motor axis, the eccentricity. The rotor 3 drives the spindle S and moves the spindle nut SM with the attached thereto or integrally formed with a piston 5, which is sealed via two seals Dl and D2 to the piston housing 7. The anti-rotation lock VS is fastened to the front end. The anti-rotation device VS has a groove 6 extending in the axial direction into which a projection V arranged at the piston end 5B2 engages and receives the motor torque Md. In addition, a radial bearing L2 is provided in the region of rotation VS. This bearing L2 is formed by the outer wall of the second piston region 5 _B2 whose outer radius d2 is greater than the outer radius d3 of the first piston region 5 _BI , in which the seals Dl and D2 are arranged. In this case, the outer wall is interrupted only by the projection V of the anti-rotation lock and is otherwise in full contact with the cylinder inner wall.

Due to the smaller diameter d3 in the region of the seals Dl and D2 prevents the piston wall of the first piston portion 5 _BI in Contact with the cylinder wall comes, so that damage to the seals Dl and D2 is safely avoided.

At higher operating pressures, the piston force acts on the ball bearing LI via the KGT helical gear. Here, the eccentricity or center deviation of the rotor 3 is compensated by the predominantly elastic design of the rotor 3, but also by bending of the spindle S, piston 5 and tilting movement of the anti-rotation lock VS. In this case, no forces act on the piston 5 and sliding surfaces of the seal D 1, D 2, since the tolerances of anti-rotation lock VS, piston 5 and cylinder bores ensure a clearance S ₀ . The anti-rotation lock VS runs in the piston 5 with little play. Accordingly, the tolerances of anti-rotation lock VS and bore are dimensioned. Thus, the reliability of the seal Dl, D2 is greatly improved, which is of particular importance in AD and FO requirements. A small lateral force acts on the anti-rotation device VS in accordance with the above-mentioned elasticity and eccentricity. In addition, a circumferential force Fu, which depends on the engine torque Md and the requirement for the operating pressure to be set in the pressure chamber A1, acts on the anti-rotation device VS.

FIG. 1 a shows a section through the cylinder bore with the anti-rotation lock VS. The rotation VS can be formed only by a groove 6 or by two, in particular opposite grooves 6, wherein in the respective groove 6 a piston 5 fixedly arranged projection V, which may be formed by a rod or a correspondingly shaped molding, intervenes.

The engine torque Md generates a circumferential force Fui on the groove 6, which is transmitted as reaction force F _u2 from the anti- _{rotation lock} VS to the cylinder bore. The rotation VS is performed over the bearing L2 in the bore with little play, for which purpose the tolerances of rotation VS and bore must be provided accordingly.

The engaging into the groove portion of the projection V can advantageously be formed slightly spherical, for example, with a radius R, so that due to the tolerances no edge pressure between the groove and projection occurs. As already stated, the rotation can "two-armed" by means of two projections V be formed, whereby the loads are derived symmetrically. However, this embodiment of the anti-rotation device may require a larger clearance s ₀ between the outer wall (d3) of the first piston area 5 _BI and the inner wall of the cylinder with its inner diameter d4.

At the beginning of a piston stroke or upon reversal of the stroke, no counterforce or axial centering force acts on the first bearing LI and / or on the ball screw transmission KGT for a short time or for a short stroke movement, which causes the rotor to tumble can lead. For this purpose, FIG. 2 shows two possible alternative solutions which, however, can also be provided jointly. In a first alternative, the electromotive drive for the helical gearbox has an additional radially acting support in the form of a bearing part 30, which is fastened to the housing 1 and extends to the end area 3e of the rotor 3 and this, at least during a tumbling motion , supported and, depending on the training with a small clearance to the rotor end 3e is arranged spaced, such that it comes into contact only with a tumbling motion of the rotor 3 with the rotor end 3e. It may also be resilient, so that no hard stop occurs. However, it can also be designed such that it always bears against the rotor end 3e.

Alternatively or in combination with the bearing part 30, a piston return spring RF can be provided which permanently exerts a restoring force on the piston 5 and thus prevents a tumbling motion.

FIG. 3 shows a further possibility for suppressing the tumble motion of the rotor 3. Here, a spring element X2, which is clamped between the housing 1 and the rotor 3, generates an axially directed force F2, which ultimately acts on the rotor 3. In this case, the spring element does not have to act directly on the rotor 3, but can also be supported on the bearing LI with its one free end in the axial direction. Between the housing 1 and the spring element X2 a bearing X3 is provided to minimize the friction. In the embodiment shown in FIG. 3, one end of the spring element X2 is clamped between the bearing LI and a sleeve X5, wherein the sleeve X5 is secured by means of a securing device. around XI on the wall portion 3W1 of the rotor 3 is secured against loss.

Due to the fact that the spring element X2 is pre-tensioned, the force F2 acts centering on the rotor 3. The force F2 can advantageously be designed such that the forces during stroke reversal or also during suction by means of the piston with negative pressure compensate. For a tumbling of the rotor is advantageously avoided. The spring element X2 may be slotted on its one side, in particular its front side, so that the corresponding elasticity is achieved without close tolerances. The load of the thrust bearing X3 is small, so that only a few balls, e.g. 3, are necessary and no high strength of the bearings is necessary. Protection against contamination or damage due to external influences can be achieved by the dashed cover X4.

The illustrated securing of the spring element X2 can, as already stated, be effected by a securing ring XI with a spacer X5. However, it is also possible for the spacer sleeve X5 to be connected to the rotor 3 by welding or calking.

The bearing L2 and the anti-rotation device VS are designed as in the embodiment shown in FIG. In this respect, reference is made to the comments on Figure 1.

FIGS. 4 to 7 show further possible embodiments in which the measures for preventing the tumbling movement of the rotor 3, which are previously shown and described in FIGS. 1 to 3, can likewise be provided.

FIG. 4 shows a construction very similar to FIG. Here, the anti-rotation VS at the same time takes over the bearing L2 in the piston bore 5z. The sliding movement runs in liquid and depending on the speed with low friction by the known hydrodynamic effect. The anti-rotation device VS consists of corresponding sliding material, preferably plastic or tin bronze. At higher operating pressures, the piston force acts on the ball bearing LI via the KGT helical gear. Here, the eccentric rizität or center deviation of the rotor 3 on the predominantly elastic design of the rotor 3, but also compensated by bending of the spindle S, piston 5 and tilting movement of the rotation VS. In this case, no forces act on the piston 5 and sliding surfaces of the seal D 1, D 2, since the tolerances of anti-rotation lock VS, piston 5 and cylinder bores ensure a clearance S ₀ . The anti-rotation VS runs in the piston 5 with little play. Accordingly, the tolerances of anti-rotation lock VS and bore dimensioned. Thus, the reliability of the seal Dl, D2 is greatly improved, which is of particular importance in AD and FO requirements. A small lateral force acts on the anti-rotation lock VS in accordance with the above-mentioned elasticity and eccentricity. In addition, a circumferential force F _Uf which depends on the engine torque Md and the requirement for the operating pressure to be set in the pressure chamber A1 acts on the anti-rotation lock VS.

In case of an error in the control or failure of the piston 5 can go back quickly, which is intercepted by a resilient stop 11, which also brakes the rotating spindle nut SM. To drive the motor and determine the piston position, the rotor 3 drives the engine sensor via gear drive, in which the shaft is connected to a target, which is preferably a Hall element and is seated on the PCB of the ECU.

The motor housing 1 is seated flat without conventional collar centering on the piston housing KG and is fastened to it by means of screws 9. For better installation, the seals and the piston raceway can be arranged in a flange piece.

During axial displacement of the piston, small hydraulic forces are applied in the initial stroke range of hydraulic applications, which are absorbed by the first bearing LI. In this case, the eccentricity of the rotating parts is also absorbed by tilting around the bearings LI and L2 or, if the anti-rotation device VS is formed accordingly. At high axial forces, the first bearing LI is preloaded so strongly that the eccentricity is absorbed by the elasticity of rotor 3, helical gears S, SM and / or adjusting element. To prevent the ingress of dirt, the ball screw, consisting of spindle S and spindle nut SM, can be provided with at least one particle scraper PS. The particle wiper or wiper PS, in particular, is intended to prevent vaporizing particles from subsequently entering the ball track during operation during assembly of the worm gear. At least a particle wiper PS, which is arranged on the open side of the ball screw drive, should be provided. This can be fixed or fixed in particular on the rotor 3, as shown, or on the piston 5. In FIG. 1, at least one of the two particle wipers PSi or PS _{2 should} therefore be provided. So that particles can not penetrate from the left through the connection rotor 3 and spindle S, a corresponding connection and / or seal can also be provided here. For safety, both particle scrapers PSi or PS _{2 can be} provided. The spindle S is located with its right end in the piston, which closes the spindle from the right hermetically. The particle wiper PSi is arranged between stop ring 11 and spindle nut SM or end face of the piston 5. The particle scraper PS ₂ is arranged between the rotor 3 and the piston 5. The particle stripper PS can advantageously be made of a material, in particular a felt-like material, whose material and dimension, in particular fibers, do not harm the ball screw drive.

Preferably, the helical gear according to the invention is mounted in a clean room.

FIG. 4a shows a section through the cylinder bore 7z with the anti-rotation lock VS. The anti-rotation lock VS acts on both sides on a segment-shaped support 6, which is here riveted to the flange piece 7, alternatively laser-welded. This solution has the advantage that the sliding surfaces in the cylinder bore 7z and the support have a high surface quality, resulting in low friction and little wear. The engine torque Md generates a circumferential force Fui on the support 6, which is transmitted as reaction force F _u2 from the anti- _{rotation lock} VS to the cylinder bore 7z. The rotation VS is performed over the bearing L2 in the hole 7z with little play, for which purpose the tolerances of rotation VS and 7z hole must be provided accordingly. sen. The engine torque Md acts in accordance with the set operating pressure significantly more in the forward movement to the pressure build-up than in the backward movement to the pressure reduction P _abl because essentially the piston force causes the return movement. The circumferential force also generates friction forces in the radial direction, which are insignificant due to the bearing of the anti-rotation device VS in the bore 7z. The radial forces are due to the elasticity, especially of the rotor 3, small. On the other hand, the axial frictional forces acting on the movement of the piston and the anti- _{rotation lock} VS are not negligible and depend on the motor torque Md and the circumferential forces Fui and F _u2 . Here, the friction coefficient is decisive. However, this is already small due to the hydrodynamics. It may additionally be reduced, be provided in the roller bearing, as shown and described in Fig. Lb.

FIG. 4b shows an alternative to the sliding guide of the anti-twist device VS. In the embodiment of Figure lb transferred gene rollers 15 and 15a, which are advantageously stored by means of needle bearings, the circumferential forces Fi and f _2. Optionally, a third roller 15b may be inserted on the opposite side of roller 15a. Again, additional bearings L2 are provided to guide the rotation VS. Also in this embodiment, the anti-rotation VS is performed over game at the bearings in the hole.

FIG. 5 shows an embodiment with a rotatable spindle nut SM. The spindle S is connected to the piston 5 or a coupling piece KO for the linear drive of a mechanism. The application is not limited to hydraulics only. Again, a rotation VS is necessary, which runs at non-hydraulic with dry bearings or grease. In such an embodiment, the seals Dl and D2 may be omitted. However, storage via bearings LI and L2 remains. In contrast to FIG. 1 with a one-piece rotor 3, a two-part rotor 14 can be used. The drive with motor is expediently connected to an ECU, for which corresponding contacts to the engine are provided. Similar to the seal arrangement of the push rod piston with primary seal D2 and secondary seal Dl volume is transported from the storage container 12 into the pressure chamber via the Schnüffelloch SL.

The particle wiper PS ₃ can be arranged or act between a carrier part 35, which is fastened or arranged on the piston housing 16 or on the flange 7, and the rotor 3 or the piston 5. Alternatively or in addition, at least one of the particle wipers PS ₄ , PS ^{'can be} provided, which can be arranged between the stop ring and the spindle nut SM or between the spindle nut SM or rotor 3 and the spindle S.

It is likewise possible to provide a closure means 36 which closes off the rotor 3 on its end face and thus prevents the penetration of protective particles from this side.

FIG. 6 shows an embodiment which essentially corresponds to the arrangement of FIG. 1, with the difference that a double-stroke piston DHK is used, which conveys volumes under pressure during forward and backward movement and seals off two pressure chambers A1 and A2 from one another separates. This is done via feed valves V, which may be simple check valves or solenoid valves. In addition, SV valves with connection to the reservoir 12 are necessary here. Of course, they are also welcome in the embodiment. Fig. 1 are used. In addition to the seals Dl and D2, another seal D3 can be provided for the stepped piston. Again, there is again a game S ₀ , so as not to burden the sealing surface. The rotation VS is arranged here in the piston 5. This consists of a profile bar 18, z. B. four- or hexagon, which supports the engine torque Md from the piston 18 via a bearing piece 19. The profile bar 18 is fixedly connected to the piston housing 16. The anti-rotation VS also forms here the storage L2.

FIG. 7 shows an embodiment which essentially corresponds to that of FIG.

2 and 3 corresponds. Here, the bearing L2 via a slide ring in the piston 5. The rotation VS is also configured here in the piston 5 as Oldham coupling. The profile bar 19 may be a rectangular profile here. The Oldham coupling piece can in principle the piston movement in the y- and z-direction balance and is floating over play S between piston 18 and profile bar 19 mounted. With this solution, the center axis of the profile bar 18 can be designed with larger tolerances.

The entire consideration of all facts, including the problems A) to G) show the complexity of the kinematics and their solution with small forces both radially and axially acting. The screw-type transmission driven by an electric motor according to the invention can advantageously be used in a broadband manner in the hydraulic system. In this case, as described above, the piston may be a single-stroke and a double-stroke piston. However, it is also possible to use the electric motor-driven screw drive according to the invention for driving a mechanism, in which case the adjusting element acts as an actuator, or clutch for a drive.

Claims

claims

1. piston-cylinder system, in particular for a control unit or brake system, with a cylinder (7) with a cylinder bore having an inner diameter (d4) and a piston (5) axially displaceably mounted therein, wherein in a first axial piston region (5 _BI ) between the outer wall of the piston (5) and the inner wall of the cylinder, at least one seal (Dl, D2) is arranged, characterized in that the piston has a second piston region (5 _B2 ), the outer diameter ( d2) is greater than the outer diameter (d3) of the first axial piston area (5 _BI ).

2. piston-cylinder system according to claim 1, characterized in that the outer diameter (d2) of the second piston portion (5 _B2 ) is formed such that the outer wall of the second piston portion (5 _B2 ) with little play in the cylinder (7), in particular with a play of 1 to 20 pm, is mounted displaceably, in particular with its outer surface over the entire circumference on the cylinder inner wall storage or rests.

3. Piston-cylinder system according to claim 1 or 2, characterized in that the second piston region (5 _B2 ) forms a full or almost vollumfängliche bearing (L2) for the piston (5).

4. Piston-cylinder system according to one of the preceding claims,

characterized in that the second piston region (5 _B2 ) is the end region of the piston (5).

5. piston-cylinder system according to claim 4, characterized in that the end face of the end portion (5 _B2 ) defines a working space (Al).

6. Piston-cylinder system according to one of the preceding claims,

characterized in that the first piston area (5 _BI ) at the two piston area (5 _B2 ) adjacent.

7. Piston-cylinder system according to one of the preceding claims,

characterized in that the outer diameter (d3) of the first piston region (5 _BI ) is formed such that even with a Verkip- pen of the piston (5) relative to the cylinder (7), the seals are not unduly stressed and / or the outer wall of the first Piston area (5 _BI ) does not come into contact with the inner wall of the cylinder bore.

8. Piston-cylinder system according to one of the preceding claims,

characterized in that the piston (5) by means of a Verdrehsicherung- (VS) is secured against rotation about its axis.

9. Piston-cylinder system according to claim 8, characterized in that the anti-rotation device (VS) in the region of the second piston region (5 _B2 ) is arranged, wherein the second piston region (5 _B2 ) rests with its circumferential surface fully on the cylinder inner wall, with Except for the area in which a projection (V) is arranged.

10. piston-cylinder system according to claim 8 or 9, characterized in that the rotation (VS) by at least one of the piston

(5) projecting projection (V) and at least one in the inner wall of the cylinder bore incorporated, extending in the axial direction, groove (6) is formed.

11. piston-cylinder system according to claim 10, characterized in that the projection has curved side surfaces (Fal), such that in particular due to tolerances no edge pressure between the groove

(6) and projection (V) occurs.

12. piston-cylinder system according to claim 10 or 11, characterized marked, that the projection (V) by a in the second piston region (5 _B2 ) arranged rod or arranged molding is formed.

13. Piston-cylinder system according to one of claims 10 to 12, characterized in that two projections (V) diametrically on the piston (5) are ordered, each engaging in a groove (6) and together the

Form anti-rotation lock (VS).

14. Electromotive driven helical gear for driving the piston (5) of a piston-cylinder system, said piston (5) by means of an electric drive (2) and a helical gear (S, SM), in particular in the form of a spindle drive, along a web, in particular axis (AX), is movable back and forth, wherein the drive (2) comprises a rotor (3) or driven by a drive transformer, by means of a first bearing (LI) in a housing (1) rotatable is mounted and fixedly connected to the input of the helical gear (S, SM) or integrally formed therewith, and that the output of the helical gear (S, SM) is connected to the piston (5) or formed integrally therewith, wherein the Verdrehserung (VS) prevents rotation of the piston (5) in the circumferential direction about the track or axis (AX), wherein

- The anti-rotation (VS) in or at the end (5 _B2 ) of the

Piston (5) is arranged or the end portion (5 _B2 ) part of

Anti-rotation lock (VS) is, wherein the end portion (5 _B2 ) is the region of the adjusting element (5) which faces away from the helical gear (S, SM),

and or

- Between the first bearing (LI) and the piston (5) of the rotor (3) or transformer, the helical gear (S, SM) and / or at least a part of the adjusting element (5) at least in a region transverse elastic to Rotationsach - Is formed or are, wherein the at least one region, in particular a spring element (FE) or an area having higher elasticity, characterized in that the piston-cylinder system is designed according to one of the preceding claims.

15. Electromotive driven helical gear according to claim 14,

characterized in that the first bearing (LI) between the rotor winding and / or permanent magnets supporting part (3p) of the rotor (3) and the input of the helical gear (S, SM) is arranged.

16. Electromotive driven helical gear according to claim 14 or 15, characterized in that between the first bearing (LI) and the end portion (5 _B2 ) and the radial bearing (L2) no further bearing for the radial support of the rotor (3) or transformer and the Schraubge- transmission (S, SM ) is available.

17. Electromotive driven helical gear according to any one of claims 14 to 17, characterized in that the rotor (3) or the transformer at least in a region between the first bearing (LI) and the input of the helical gear (S, SM) transversely elastic is formed to the axis of rotation, in particular a spring element (FE) or a region having higher elasticity, which is in particular greater than the elasticity or deviation of the rotating parts of the central axis (AX).

18. Electromotive driven helical gear according to one of the preceding claims, characterized in that between the helical gear (S, SM) and the adjusting a spring element or a resilient region is arranged, whose elasticity is greater than the elasticity or deviation of the rotating parts from the center axis (AX) and / or that the screw thread (S, SM) is elastic or flexible transverse to its axis of rotation (AX).

19. An electromotively driven helical drive according to any one of claims 14 to 18, characterized in that either the input of the helical gear (S, SM) the spindle nut (SM) and the output of the helical gear (S, SM) the spindle (S)

or

the input of the helical gear (S, SM) is the spindle (S) and the output of the helical gear (S, SM) is the spindle nut (SM).

20. An electromotively driven helical drive according to one of claims 14 to 19, characterized in that the rotor (3) or transmitter at least in the range from the first bearing (LI) to the input of the helical gear (S, SM) in Cross-section is double-walled, wherein the two walls (3 _wi , 3 _w2 ) via a curved, in particular in cross-section U-shaped, wall portion (3 _w3 ) with each other are connected, in particular integrally formed, welded together, riveted, crimped or glued.

21. Electromotive driven helical gear according to claim 14,

characterized in that the end portion (5e), a form-fitting part (22) and the engagement part (19) form an Oldham coupling (Figure 4a).

22. An electromotively driven helical drive according to one of the preceding claims, characterized in that the outer diameter (d3) of the first piston region (5 _BI ) is smaller than the inner diameter (d4) of the cylinder (7) by a clearance (s ₀ ). in particular the clearance (s ₀ ) is between 0.1 mm to 5 mm, preferably 0.2 to 0.8 mm.

23. Electromotive driven helical gear according to one of claims 14 to 22, characterized in that a channel between the seals (Dl, D2) opens in the cylinder interior, which is in communication with a storage container (VB).

24. Electromotive driven helical gear according to any one of claims 14 to 23, characterized in that by means of the piston (5) in at least one pressure chamber or associated therewith or connected hydraulic circuit or circuits held a pressure up and / or can be degraded, in particular that the adjusted or regulated pressure is used to adjust the wheel brakes, clutches and / or gear actuators.

25. Electromotive driven helical gear according to the preamble of claim 14 or any one of claims 14 to 24, character- ized in that a radial support (30) for the end (3e) of the rotor (3) is provided, which on the housing (1) is arranged and on the end (3e) of the rotor (3), in particular at the beginning of the stroke of the piston (5) or when reversing the direction of the stroke acts

and or

a device (X2, RF) is provided for exerting a force (F2) or restoring force acting on the rotor (3) and / or the piston (5) in the axial direction.

26. Electromotive driven helical gear according to claim 25, characterized in that the end (3 e) of the rotor (3) on the bearing (LI) facing away from the rotor stator (RS) and / or the support (30) with a game is spaced from the rotor end (3e) and comes into contact only with a tumbling motion of the rotor (3).

27. Electromotive driven helical gear according to claim 25,

characterized in that by means of a force acting on the piston (5) return spring (RF), the restoring force is generated.

28. Electromotive driven helical gear according to claim 25,

characterized in that the device comprises a spring element (X2), wherein the spring element (X2) on the one hand on the housing (1) and on the other hand on a rotating part, in particular in the form of the rotor (3), bearing (LI), one on the rotor (3) or between a rotor (3) and the helical gear (S, SM) arranged part supported.

29. Electromotive driven helical gear according to claim 28,

characterized in that between the spring element (X2) and the housing (1) a bearing (X3), in particular in the form of a ball bearing, is arranged.