DE19633412A1 - Pressure operated drive - Google Patents

Abstract

The drive consists basically of the working cylinder (1), a control transmission (2) and an axial shell bearing (11). These groups are held in position relative to each other by mounting components (7, 8, 12). They can be designed for integration with a single-axis cylinder. The control of the piston (3) and its rod (4) is effected by a non-self-locking threaded spindle (9) carrying a spindle nut (10) connected to the piston, which can be activated proportionally by a control shaft (15) connected to it. The actuator moments can be set as required by switchable control pressure connections (20).

Description

The following publications have been considered for the preparation of the description:
DE 33 11 307 A1
DE 41 41 460 C2.

The invention relates to a pressure medium-operated one-sided or Working cylinder closed on both sides with cylinder cover circular, rectangular, square, oval or one any conveniently shaped work area cross-section, one axially in this slidably guided working piston and the working piston assigned coupling elements, for example with the working piston connected piston rods, bands, ropes or permanent magnetic Elements that allow the axial displacement and the forces of the To tap the working piston positively or non-positively.

Working cylinders of the type mentioned above are commercially available and wide widely used. Different designs of the working cylinders in Reference to the assigned coupling elements are:

• 1. The so-called double-acting cylinder with two connections for supplying and discharging the pressure medium into the workrooms to both Sides of the working piston and one-sided from one of the two Piston rod leading out of cylinder cover or with on both sides Piston rod led out from both cylinder covers as a coupling element.  
• 2. The so-called single-acting cylinder. The design is similar to the double-acting described in item 1 Working cylinder, but with only one working area on one side of the Working piston and only one connection for feeding and discharging the Pressure medium. The pressure medium can only work in one piston Move direction, the return movement is done either by a spring, often built into the working cylinder, or by a retroactive one Force on the piston rod. The coupling element is also one piston rod on one side or on both sides.
• 3. The so-called edge or rope working cylinder. This design is partly the double-acting described in item 1 Working cylinder with piston rods led out on both sides similar. Instead of the piston rods there is a band or rope working cylinder flexible band or rope arranged, but appropriate, in the Arranged cylinder caps, the rope or tape is made of seals led out of the working cylinder and over on the cylinder covers attached deflection rollers with one arranged on a guide Connected carriage, which in this case is the coupling element.
• 4. The so-called cylinder without piston rod. Here is the strength of the working piston is not tapped by coupling elements which exit through the cylinder covers that close the work area, but the force of the working piston is parallel to the axis of the Working cylinder next to the working cylinder is positive or non-positive tapped. Coupling element is an outside on the working cylinder arranged and axially parallel slide, the form or is non-positively connected to the working piston. The adhesion takes place via magnetic forces through a cooperation of Permanent magnets and ferromagnetic components, each in the Working piston and the carriage are arranged. In the positive locking are pistons and slides by one Longitudinal slot in the working cylinder mechanically connected to each other, the longitudinal slot in the working cylinder from the inside through a suitable sealing tape is closed. In the middle of Working piston, between those on the front of the working piston attached sealing sleeves, the sealing tape is through on the carriage  attached runners bent in the direction of the piston axis and thus arise between the slot edges of the working cylinder and the sealing tape Openings for the mechanical connection of the working piston with the Carriage.
• 5. The so-called rotary or swivel cylinder. With this type of Working cylinders come to different operating principles Application. With the inventive concept described below only the functions described below are of importance. The linear movement of the working piston acts via a non-rotatable Spindle nut on a rotating non-self-locking Threaded spindle, with one, from the working cylinder led out, drive shaft is connected as a coupling element of a swivel or rotary movement can be removed. Known is also a functional principle by a kinematic reversal is achieved, with a non-rotatable threaded spindle rotatable mounted spindle nut drives.

The above-described pressure-operated working cylinders have the following defects due to the system:
The precision of the work movements derived from the movements of the working piston, such as linear, swivel or rotary movements, is mutually dependent, on the one hand on the specific properties of the pressure medium such as compressibility, volume expansion coefficient, viscosity-temperature behavior, lubrication properties and on the other hand on the leakage losses, of pressure medium generation and pressure medium control. For some areas of application, this mutual dependency results in an undesirable unstable, poorly reproducible operating behavior of the working cylinders, because the specific properties mentioned above fluctuate too much under operating conditions. This applies in particular to pneumatic working cylinders. In addition, leakage losses are unavoidable and therefore loads or retroactive forces cannot be kept motionless without complex regulation. With known brake devices on the piston rod, the disadvantage can only be remedied insufficiently, because after the brake is released, the previously braked movement is released, ie an undesired piston movement can take place. Very small adjustment paths and adjustment speeds can often not be achieved even with complex control devices, because the compressibility of the pressure medium and the friction on the sealing elements of the working piston lead to so-called stick-slip effects and this leads to uncontrolled jerky jumps. In the case of pneumatic working cylinders, an undesired venting of one side of the piston can result in an explosive piston displacement, possibly with destructive consequences and risk of accidents. The synchronization of two or more cylinders, for example for lifting a platform, can only be implemented very cost-intensively.

The invention is based, pressure medium-operated task Working cylinders of the generic type in such a way that the systemic defects described above are avoided and Working piston travel with great precision and a high degree Controlled in reliability and stop positions independent of pressure medium can be fixed. Especially the rather problematic one kinematic and dynamic behavior of pneumatic cylinders is to be improved by the invention.

This task is the case with pressure-operated working cylinders Generic type according to the invention by the characterizing features of claim 1 solved. Advantageous embodiments of the Invention are presented in the subclaims.

The advantage of the invention is the fact that in many Variations known cylinder is improved so that it for Areas of application can be used, which so far only with very complex drives could be realized. The achieved The advantages are that a movement of the Working piston can only take place if the self-locking in the Control gear with a control rotation, generated by one Control drive and introduced into the control shaft of the threaded spindle, is overcome. Another advantage is that the travel of the working piston is proportional to the angle of rotation of the control rotation  and at the same time mechanical reliable self-locking is given, the very low elastic in every stop position Behavior of a threaded spindle subjected to tension or pressure owns.

The control torque for generating a control rotation can be very small be when the efficiency of the control gear in the Power flow direction, based on the force of the Working piston on the spindle nut and further over the Threaded spindle and the one assigned to it, if applicable Thrust bearing up to the control shaft, zero or slightly less than Is zero. As you know, self-locking results in a negative Efficiency and means blocking an inevitable Mechanism or transmission solely by such frictional forces, which the external driving force, here the force of the working piston, themselves are caused and these are proportional. The required control torque depends on the force of the Working piston and the absolute size of the negative efficiency, that is, a correspondingly small absolute size of the negative Efficiency also has a correspondingly small control torque Episode. Theoretically, this would also be the case with zero efficiency required control torque is zero, regardless of the force of the Working piston. Noteworthy is the fact that a steering turn to overcome self-locking to an efficiency equal to one leads, based on the force of the Working piston and the force removable from the working piston, that is, the pressure and tensile force of a working fluid operated cylinder is not influenced by the training according to the invention, So not reduced, compared to a cylinder more well known Type, without the training according to the invention.

The situation described above shows another advantage of Invention the remarkably easy to implement property of a Power amplifier on the constructive design of the Control gear with an efficiency of zero or slightly smaller than zero, gives a very large gain.  

Are made in a further advantageous embodiment of the invention the direction of the control torque or the control rotation signals derived to control the pressure medium for the simultaneous Actuation and depressurization of the corresponding work rooms in the working cylinder, or when switching off the Control drive the depressurization of both workrooms, so are with the direction of rotation of the control rotation automatically travel in two Realize directions, i.e. back and forth. Yourself Depressurization for both workrooms has no influence on the Holding force and the set holding position. The fact that a Working piston even without pressure medium once set Hold position and with regard to security and stability with is comparable to a highly loaded screw drive Working cylinder according to the invention to a versatile Construction component, including for mobile devices such as for example mobile cranes, construction cranes or earth-moving machines for example, extending support legs that last for hours or even Weeks to fix a stand position.

How the control rotation is generated appropriately depends on the Task starting with a working cylinder according to the invention should be solved. Because the tax performance compared to that of Working piston tapped power, is very small, arise extremely versatile and in particular also very simple and inexpensive ways to apply the control rotation, such as Example of initiating a control turn by hand, if necessary by interposing a flexible shaft to connect with the Working cylinder to create a remote control. With several flexible bends driven by the same gear Waves, is a constant path adjustment of several working cylinders too to reach. So that's the problem, for example, with four Working cylinders the plane-parallel adjustment of lifts or whole assemblies in machines, very inexpensive to solve. A a similar solution can also be achieved with timing belts or chains, which drives the control turns for several Generate working cylinder. With several motors rotating in the same direction a constant adjustment of a corresponding number of working cylinders  to be achieved in the same way. Also unequal, selectable travel ranges Several working cylinders can be operated with computer-controlled motors as a control drive, for example for working cylinders as Drive for dosing pumps for different precisely to be observed Mixing ratios. While it is next to the very simple one Manual operation, in principle, every motor that generates a rotation applicable. With electric motors because of the variety of designs and the control equipment it offers pressure-operated working cylinders according to the invention Areas of application opened up for the well-known working cylinder generic type could not be used so far.

In particular, computer-controlled electromotive control drives are suitable in cooperation with analog or digital encoders that excellent and proven properties of these drives on the Quality of the movements that can be derived from the working cylinder transfer. In summary it can be said that in that only to apply a very small control output for a control drive is for pressure actuated working cylinders of any type and size new areas of application are opened up.

An advantageous embodiment of the invention results from the fact that in the control gear of a non-self-locking threaded spindle, for Example of a ball screw, a self-locking one Thrust bearing is assigned, which is due to the thread axial force is self-locking, the axial lock bearing according to the invention by a constructively simple design is realized in the way that in two preferably fixed at a distance from each other identical abutments, each facing each other opposite thread surfaces a thread-free shaft part of the Threaded spindle is mounted and firmly connected to it between the Load surfaces an axial disc is arranged, which axial axial axial forces across the end faces the thrust washer on the corresponding load surface of an abutment transmits, and one for the thread axial forces in two directions represents effective thrust plain bearing, and by friction on the rotation the threaded spindle generates counteracting inhibiting torque that depends is of the axial force, the direction of the axial force and the effective one  Friction diameter of the axial disc and the coefficient of friction of the Material pairing. The dimensions of the axial washer are expedient chosen so that the distance between the end faces by a small amount is smaller than the distance between the load surfaces and the outer diameter so large that a sufficiently large effective friction diameter arises for an inhibiting moment, to generate a negative efficiency in the desired size. According to the invention results from the constructive Connection of a non-self-locking threaded spindle with a Axial inhibitor bearing the advantage that the, a screw drive own normally non-variable efficiency with the thrust bearing can be adapted specifically to the respective application, in particular, an interpretation close to a desired one is very simple Efficiency zero possible to get a small control torque.

However, the effective friction diameter can only be insufficiently accurate can be defined if the entire end face of the axial washer is defined as Frictional surface acts on the load surface, in particular can by running time conditional wear change the effective friction diameter. According to the invention, it is therefore advantageous on the end faces Arrange recesses centrally on both sides of the axial disk, which in each case from the center of the end faces to almost the outer edge of the Front faces are sufficient so that only narrow ring faces rub against can act on the flat load surfaces of the abutments, and thus the effective friction diameter pretty much in the middle of the The end face of the ring can be assumed to be horizontal and so also with one wear caused by running time a change in the effective Friction diameter does not occur. The ring faces described above can also be realized constructively by friction rings, which are in suitable recesses on both sides on the face of the axial washer be pushed in and protrude slightly from the front.

It does not matter whether the ring faces are part of the axial disk or caused by pressed-in friction rings, in any case According to the invention, the ring end faces result in more Advantages such that when supporting an annular end face a closed space is created on a load surface Pressurizable, a relieving the ring face  Axialdruckkammer represents which is radially limited by the Ring face and the shaft part entering the load surface and is axially limited by the load area and the opposite of the End face of the axial disc slightly recessed. Is through pressure holes in the abutments in the respective closed axial pressure chamber introduced, so creates Compensation force that opposes the thread axial force is directed, and thus the inhibiting moment is reduced. With a Pressure control of the pressure medium introduced in the axial pressure chamber is an extremely sensitive, fast and accurate according to the invention Regulation of the inhibiting torque possible because the only active element of the Pressure of the pressure medium in the axial pressure chamber is a regulation of the inhibiting torque is given without friction and hysteresis. Furthermore can the inhibiting torque by appropriate pressure in the axial pressure chamber also be lifted to the extent that the non-self-inhibiting Threaded spindle turns automatically due to the force of the working piston, where the usable piston force from zero to a maximum size adjusted with the pressure in the axial pressure chamber respectively can be regulated continuously. The same applies analogously to that at the Torque acting torque, which is now on the control shaft Downforce can be removed. With one of the control shafts assigned, depending on the speed of rotation, rotary damper is in addition to the adjustable piston forces and / or torques created the possibility of the speed of the working piston or the rotational speed of the control shaft corresponding to that effective output torque on the control shaft continuously change, that is, increased with increasing output torque the rotation speed of the control shaft. That from the Rotational speed is dependent principle of action of the rotary damper the tough liquid damping and / or the overcoming of the shear forces a viscous liquid, such as those from the Motor vehicle technology known viscous clutches. As another The advantage according to the invention results from the sealing effect of the Ring face when rubbing due to slightly flowing pressure medium lubrication and cooling of the load surfaces.  

The above-described adjustable thrust bearing has particular advantages with an adjustment of the efficiency, because of a possible small control torque zero or slightly less than zero should. The coefficients of friction that determine the efficiency are both in the Thread as well as in the thrust bearing under operating conditions, however not constant. So are the static friction values for physical reasons inevitably greater than the sliding friction values, that is, for starting of the working piston from standstill is a larger control torque required to maintain movement. An interpretation of the control gear with an efficiency close to zero in order to To get by with the smallest possible control torque are limits set. On the one hand there is a big negative at standstill Efficiency desirable because it increases the safety of the Self-locking is increased, so there is also a large coefficient of static friction Cheap; on the other hand, efficiency is close to zero aimed to get by with small tax moments. This Discrepancy can be resolved by designing the thrust bearing with a larger self-locking reserve Friction diameter in connection with regulating devices and / or Control the pressure medium for the application of each closed axial pressure chamber.

According to the invention, the two solutions described below are proposed for the device for regulating and or controlling the pressure medium:
A first solution results from the fact that each control pressure connection to an axial pressure chamber is assigned a pressure increase valve and a pressure outlet valve, each valve having two switch positions "open" and "closed". The valves can be identical in construction or combined to form a valve element with the same effect. A pressure medium distributor supplies both valves via adjustable reducing valves with different inlet pressures, the pressure increasing valve being supplied with the higher pressure than the pressure outlet valve.

It is assumed that a, with a control gear according to the invention connected, double-acting cylinder when initiating or Switching on a control torque to generate a control rotation via a simultaneously directionally compatible with respect to the direction of the Control torque switching filling valve the corresponding Working space of the working cylinder is filled with pressure medium or is already filled in advance, so with the valve arrangement described following function according to the invention can be realized.

Is when initiating or switching on the control torque - if necessary in addition to the filling valve - at the same time corresponding pressure increase valve to the "open" position switched, which for example by a in the Pressure medium supply arranged throttle acts such that a continuous increase in pressure of the pressure medium in each closed axial pressure chamber, this is reduced Inhibiting torque and falls below the control torque after a short time, that is, the lead screw starts rotating, that now only the smaller sliding friction coefficient takes effect, including that possibly existing and desired in the static friction coefficient Self-locking is no longer given and the threaded spindle by itself would keep turning. Are signals from the beginning of the rotation to Switch the pressure increase valve to the "Closed" position and the Pressure outlet valves derived from the "open" position, which results in a Reduction to a previously set lower pressure takes place, see above in the axial pressure chamber a pressure-proportional, now smaller one is created Compensation force, which the momentarily smaller inhibiting moment at Transition to the coefficient of sliding friction is now increased again, so that the Self-locking is guaranteed again and a rotation of the Threaded spindle can only take place if the previously existing Control torque acts on the threaded spindle. That is, that The control torque is approximately the same despite different friction values. If the entire system is depressurized, there is also every relief in the axial pressure chamber so that the ring face of the Axial disk again generated the desired self-locking reserves.  

As a second solution according to the invention for controlling the Pressure medium is proposed, the control shaft of the threaded spindle at least one torque measuring and the direction of the torque assigning detecting device, the device a Evaluation circuit is assigned, which depends on the recorded values representative analog and / or digital control signals generated that act on pressure valves, which are connected to the Abutments are connected, and depending on the direction of the torque apply pressure medium to each closed axial pressure chamber, wherein the pressure valves control the pressure in steps or continuously regulate proportionally in such a way that with increasing size of the Control torque by increasing the pressure in the axial pressure chamber Reduction of the inhibiting torque occurs in the thrust bearing, and thus the initially greater static friction between the spindle nut and threaded spindle and the axially loaded load areas in the thrust bearing by corresponding reduction of the inhibiting torque by the control torque is overcome. The now after overcoming static friction adjusting sliding friction leads to a reduction of the inhibiting torque, synonymous with a reduction in the control torque, synonymous with a reduction in pressure in the Axial pressure chamber and ultimately approximately to an equilibrium between control torque and inhibition torque, which is automatically the adjusts the existing control torque.

The above-described facts of the invention as a whole the principle of self-locking. Overcoming self-locking an initiated control rotation does not change anything Effectiveness of self-locking when the control rotation is switched off causes the piston movement to stop suddenly, if a certain so-called critical mass ratio between the working mass and the tax mass is fallen below, the Working mass the sum of the masses to be assigned to the working piston represent and the control mass the sum of the powertrain with the Representation of control rotation related masses. A jerky Stopping the working mass can cause considerable retarding forces be connected and with a mechanical destruction of the Drive system.  

This disadvantage can be a further embodiment of the invention thereby be prevented that arranged in the control gear Axial brake bearings as axial brake bearings or axial shift brake bearings is trained. The axial brake bearings and axial shift brake bearings are in Sub-areas functionally identical to that already described Axial inhibitor bearing, here also one with ring end faces works Axial washer on opposite load surfaces, including the Function and design of the axial pressure chambers is also identical. Deviating from the described construction and function of the Thrust bearings are in the thrust bearing and thrust brake bearing the load areas are arranged on rotating disks, which are in separate Support bearings, preferably in roller bearings, are rotatably mounted, and the axes of rotation of the rotating disks to that on the threaded spindle attached axial disc are axially aligned with the rotating discs at least one, on both rotating disks or only one at a time Acting brake is assigned to the turntable. As long as the Co-rotating disc, held by the brake, not rotating, is the Axial disc in cooperation with those on the rotating disc arranged load surfaces functionally the same as the simple thrust bearing. Exceeds the self-locking when the control rotation is switched off If the braking torque on a load surface, the torque turns With the turntable, but is still braked in a defined manner, so Overloading due to excessive inhibition of self-locking does not possible. Nevertheless, the self-locking function is retained, however only up to the limit value of the braking torque.

In one embodiment of the invention, the turntable Non-rotatably axially movable brake rings assigned by Spring force in the same axial direction of force on specially for this designed braking surfaces of the turntables act like the respective thread axial forces of the threaded spindle, the Braking surfaces both perpendicular and flat to the direction of the braking force can also be inclined, ie forming a cone. in the the latter case is due to the tapered design of the braking surfaces same axial force increases the surface pressure in the braking surfaces, thus generates a larger braking torque or a predetermined one  Braking torque is to be generated with a smaller axial force. That the Brake pressure forces on the rotating disks in the same direction act like the respective axial force of the threaded spindle, in this respect advantageous that a bearing play in the rolling bearings of the rotating disks when changing the direction of the axial thread force of the threaded spindle no change in the load direction in the rolling bearings of the rotating disks causes and thus a, the life of the rolling bearings, favorable bias is given. At the same time it results thereby reducing the reversal margin of the timing gear Advantage of the use of preloaded ball roller drives in Control gear can be of considerable importance.

Axial brake bearings and axial shift brake bearings are related to the Functions previously described for this purpose the same Axial switching brake bearing also has a device for Brake release. The brake of the rotating disks can be switched off designed. When the brake is switched off, the applied one turns Turntable with almost no resistance, that is, a torque a non-self-locking threaded spindle drives over the frictional connection the thrust washer, the bearing of the With the turntable also an easily rotating bearing for the Axial forces of the threaded spindle. A positioning of the Through the working piston at any point on the commute However, switching on the brake is also possible here, the Torque-limiting brake is also important to the Switching on the brake then self-locking only up to one predetermined braking torque to take effect, otherwise Overloads due to excessive deceleration forces cannot be excluded are. Becomes a working cylinder without a control drive only through Turn off self-locking with a non-self-locking Operated threaded spindle in the control gear, it can according to the invention to avoid impermissibly high speeds of the working piston be expedient, the control shaft at least one rotary damper assign, to limit the speed of rotation, its operating principle the tough liquid damping and / or the overcoming of the shear forces is a viscous liquid, such as those from the Motor vehicle technology known viscous clutches. With several  switchable rotary dampers or at least one controllable Rotation damper is gradual or infinitely variable in speed of the working piston adjustable. There is a rotation damper According to the invention in the simplest case from one, with a highly viscous Liquid-filled and appropriately sealed roller bearing, the Inner rings rotatably connected to the control shaft and its Outer rings are rotatably arranged in the control gear. Others too Elements that can pump liquids through throttles, for example come into question as a rotation damper.

An obvious further advantageous embodiment of the invention arises for use in rotating and swiveling movements. It is known, from a linear movement of pressure-operated piston with non-self-locking screw drives to rotate and swivel produce. However, the effective screw drive in this context has nothing with the screw drive mentioned in connection with the invention to do. If the linear movement of the Working piston in a rotating or swiveling movement with known Constructions implemented, so arise analogously for the turning and Panning movements the same control options and precise Movements as described above for the linear movements. Because the mass forces during rotary movements - which may come from can exist several revolutions - especially with pneumatic Drives can reach significantly large values inventive mastery of the speed of the Linear motion, which is converted into a rotary motion, special advantages to avoid unwanted high turning speeds prevent.

Embodiments of the invention are shown in the drawings and described in more detail below.

Fig. 1 schematically shows a double-acting working cylinder in cooperation with a control gear which Axialhemmlager and consists of a non self-locking threaded spindle,

Fig. 2 shows a working cylinder with built-in control gear,

FIG. 3 shows further configurations of the control gear in relation to the axial locking bearing shown in FIG. 2, which in FIG. 3 shows a further configuration of the axial locking bearing to form a so-called axial shift brake bearing. The largely similar but somewhat simpler axial brake bearing is also described with reference to FIG. 3.

Fig. 1 shows a more schematic arrangement of a double-acting cylinder 1 of known type to which the control gear 2 is attached. The following description dispenses with the constructional details of double-acting working cylinders due to the known commercial design of the working cylinder 1 . The working cylinder 1 in FIG. 1 is an example of pressure medium-operated drives, of whatever type, whereby in each case the working piston 3 is connected to a control gear 2 and the connection is of the type or similar or equivalent, as described below.

This simple representation was chosen to be according to the invention amazingly simple, with few components, realizable Control option for pressure-operated working cylinders to show.

On the piston rod 4 , an output connection 5 is arranged, as is usually attached in various embodiments to piston rods of working cylinders of a known type. At the end of the piston rod 4 , a bushing 6 is axially fixed in front of the output connection 5 , but is rotatably arranged. A connecting web 7 which is fixedly connected to the bushing 6 is the force-transmitting link to the control gear 2 .

The control gear 2 consists of a, in relation to the working cylinder 1 , stationary frame 8 , in which a threaded spindle 9 is screwed radially rotatably connected with a spindle nut 10 arranged thereon, and the axial bearing for the direction changing thread axial forces in one of the threaded spindle 9 so-called axial check bearing 11 takes place. Further details of the design and function of the axial check bearing 11 result from the following description.

The control gear 2 is connected to the working cylinder 1 via the frame part 12 such that a threaded spindle 9 arranged in the control gear 2 is axially parallel to the working cylinder 1 .

The connecting web 7 is fixedly connected to the spindle nut 10 arranged in the control gear 2 on a threaded spindle 9 and transmits the forces and movements of the working piston 3 to the rotationally fixed spindle nut 10 , which introduces a spindle torque into the non-self-locking threaded spindle 9 . A rotation of the threaded spindle 9 is prevented by the axial locking bearing 11 , however, because the axial locking bearing 11 is designed so that the force-proportional inhibitory torque generated on the load surfaces 13 a and 13 b by the friction rings 14 a and 14 b in cooperation with the axial force is always somewhat independent of the load is greater than the spindle torque introduced into the threaded spindle 9 by the spindle nut 10 . The shaft end emerging from the axial locking bearing 11 serves as a control shaft 15 , into which a control torque which is slightly greater than the difference between the inhibiting torque and the spindle torque must be introduced in order to overcome the self-locking.

Axialhemmlager 11 consists of two stationary in the control gear 2 , with load surfaces 13 a and 13 b provided abutments 16 a and 16 b, the load surfaces 13 a and 13 b facing each other with a distance, and an axial disc 17 arranged therebetween . The axial disk 17 is arranged on a thread-free shaft part 18 of the threaded spindle 9 and is firmly connected to the latter. The thread-free shaft part 18 extends radially supported through both abutments 16 a and 16 b with one shaft end emerging as a control shaft 15 . The thrust washer 17 is designed such that cylindrical recesses 17 are arranged centrally on both end faces of the thrust washer, into which, somewhat above the end faces, friction rings 14 a and 14 b are pressed in in a rotationally fixed manner, so that the friction rings 14 a and 14 b face the ring form the axial disk 17 and, in the case of axial loading, only the ring end faces of the friction rings 14 a and 14 b have contact with the load faces 13 a and 13 b, the axial distance between the ring end faces being smaller by a small amount than the distance between the load faces 13 a and 13 b . In the event of an axial load, the axial disks 17 provided with friction rings 14 a and 14 b form so-called axial pressure chambers 19 a or 19 b with the respective load surface 13 a or 13 b, which are arranged by control pressure connections 20 a and 20 b arranged in the abutments 16 a and 16 b Pressure medium can be applied, and so reduce the specific surface pressure between the ring end faces and the load surface 13 a and 13 b, that is, the inhibiting torque is correspondingly smaller. To avoid leakage of the pressure medium, which can escape through the bearing gaps of the shaft part 18 in the abutments 16 a and 16 b, or to keep them low, the bearing gaps are selected to be very small and long, or the shaft part 18 is included in the abutments 16 a and 16 b arranged shaft seals (not shown) sealed.

Fig. 2 shows an operating cylinder 32 with an integrated control gear, said elements of the control mechanism with components of the working cylinder 32 form functional units. Therefore, the control gear does not appear explicitly in the description of FIG. 2.

Fig. 2 shows a cylinder cover 30 and once with a Axialhemmlager 31 sealed working cylinder 32, with a guided therein working piston 33, and with this non-rotatably connected a hollow cylindrical, closed at one end and sealed with the side of one side of the cylinder cover 30 extending out a fluid-operated both sides once Piston rod 34 . Extending into the piston rod 34 there is a rotatable and non-self-locking threaded spindle 35 which is fixed in the axial direction with respect to the working cylinder 32 in an axial locking bearing 31 and which is screwed into a spindle nut 36 which is firmly connected to the working piston 33 . The Axialhemmlager 31 , which consists of two identical abutments 37 a and 37 b and an axial disc 38 arranged between the abutments 37 a and 37 b, has the functions of sealing the working space 39 of the working cylinder 32 similar to a cylinder cover, the thread-free shaft part 40 of the threaded shaft 35 fluid-tightly led out of the working cylinder 32, to fix the lead screw 35 axially and to generate a self-locking of the threaded spindle 35 restraining torque, which is not self-locking adjustable by the pressure medium within the limits of self-locking up.

The following description refers because of the identical configuration of the two belonging to a Axialhemmlager 31, abutment 37 a and 37 b on only the abutment 37 a. In Fig. 2 elements of the abutment 37 a described with a number and the letter "a" are named, a symmetrically arranged and identical element is identified in Fig. 2 with the same number and the letter "b", z. B. load area 45 b.

The spatial shape of the abutment 37 a is that of a rotating body. The abutment 37 a, which is provided with a centrally arranged, stepped, continuous central bore 41 a, consists of the support cylinder 42 a with the largest diameter and connected to this, a cylinder shoulder 43 a stepped flat on one side. The axial length of the abutment 37 a is limited on one side by the flat surface 44 a on the cylinder shoulder 43 a and on the opposite side by the flat surface on the support cylinder 42 a, the so-called load surface 45 a. The central bore 41 a has the task of storing the thread-free shaft part 40 of the threaded spindle 35 radially and pressure-tight and to create a pressure medium line for the pressure medium emerging in the load area 45 a. For this purpose, part of the central bore 41 a serving for the storage and sealing of the shaft part 40 is somewhat enlarged from the load surface 45 a, so that an annular space 46 a is formed between the shaft part 40 and the central bore 41 a as a flow channel for the pressure medium. The annular space 46 a extends in the abutment 37 a to a connection to a radial bore 47 a arranged leading radially outwards in the support cylinder 42 a, which leads further to a control pressure connection 48 a arranged in the cylinder tube in the cylinder tube 51 .

The two structurally identical abutments 37 a and 37 b face each other with the load surfaces 45 a and 45 b with a spacer ring 49 between the load surfaces 45 a and 45 b in the same diameter as the support cylinders 42 a and 42 b, the cylinder tube diameter slightly increasing diameter recess 50 des Cylinder tube 51 non-rotatably and fluid pressure tight. The components axially one behind the other, the support cylinder 42 b, the spacer ring 49 and the support cylinder 42 a, are axially fixed by a precisely fitting arrangement between the end edge 52 of the diameter recess 50 and a retaining ring 53 , which is blown into a groove provided in the diameter recess 50 . In the cylindrical cavity 54 , the axial disk 38 is arranged, which is firmly connected to a thread-free shaft part 40 of the threaded spindle 35 , which passes through the central bores 41 a and 41 b and emerges from the abutment 37 a or working cylinder 32 , the from the Abutment 37 a emerging shaft part represents the control shaft 55 .

The axial disk 38 has on both sides on the end faces 56 a and 56 b projecting into the lateral surface of the outer diameter ring-cylindrical recesses, in which friction rings 57 a and 57 b are arranged, with their annular end faces slightly above the plane of the end faces 56 a and 56 b of the axial disk 38 protrude so that only the ring end faces of the friction rings 57 a and 57 b come into contact with the corresponding load surfaces 45 a and 45 b, and at the same time together with the load surfaces 45 a and 45 b each an axial pressure chamber 58 a or 58 b form. The pressure medium is fed into the respective axial pressure chamber 58 a or 58 b through the annular space 46 a.

The function of the devices according to the invention according to FIGS. 1 and 2 is explained in more detail below. It should be noted that the functions are identical in both representations, the difference is only to be seen in a different arrangement of the functional elements.

It is assumed that the working space 21 or 39 is pressurized via the pressure medium connection 22 or 59 . The spindle nut 10 or 36 transmits the piston force in the direction 23 or 60 to the threaded spindle 9 or 35 , into which a spindle torque is thereby introduced, but which does not cause rotation, because at the same time the friction ring 14 a or 57 a of the axial disk 17 or 38 is pressed onto the load surface 13 a or 45 a of the abutment 16 a or 37 a. The reaction force in the abutment 16 a or 37 a, with the effective friction diameter of the friction ring 14 a or 57 a, causes an inhibiting torque which hinders the rotation, which is calculated as follows: friction radius times reaction force times coefficient of friction. As can be seen, for a given reaction force and a given coefficient of friction, an inhibitory torque of any size can be structurally determined by choosing an appropriate friction radius. The axial pressure chamber 19 a or 58 a, which can be filled with pressure medium, allows the reaction force of the abutment 16 a or 37 a to be divided into a part subject to friction and into a smooth part, the friction ring 14 a or 57 a, the part subject to friction and the part filled with pressure medium Part of the axial pressure chamber 19 a and 58 a represent the smooth part. It should be noted that the smooth part actually represents a type of hydrostatic or aerostatic bearing which is not entirely smooth, but whose coefficient of friction in connection with the invention is so insignificant that the coefficient of friction can be assumed to be zero.

If a large friction radius is selected by means of a correspondingly large friction ring 14 a or 57 a, self-locking can be achieved for the threaded spindle 9 or 35 , and moreover a high degree of security of the self-locking. However, this can be changed by introducing pressure medium into the axial pressure chamber 19 a or 58 a and by adapting the pressure to the desired operating conditions down to the area of non-self-locking. A movement of the working piston 3 or 33 can be very slow and sensitive in connection with a control torque introduced into the control shaft 15 or 55 and / or by the pressure in the axial pressure chamber 19 a or 58 a, depending on the size of the inhibiting torque, i.e. also with low force, or can also be moved at high speed when self-locking is canceled. When the piston force is reversed, the larger piston area results in a greater piston force at the same pressure of the pressure medium, but this produces the same effects due to the functional proportionality in the symmetrical axial thrust bearing 11 or 31 . At most, a larger control torque may be required, but this can be reduced again by a correspondingly higher pressure in the axial pressure chamber 19 a or 58 a.

. Figure 3 shows the Axialschaltbremslager a control mechanism and it derived - as will be explained - a broadly similar but somewhat more simply constructed Axialbremslager. First, an axial shift brake bearing of a control gear is shown, to which a non-self-locking threaded spindle 70 is assigned. The axial shift brake bearing is integrated in a cylinder tube 71 of a working cylinder 72 as an example. However, the Axialschaltbremslager described in the FIG. 3 can find application also 11 or 31 in place of the Axialhemmlager described in FIG. 1 and FIG. 2. In the illustration, the working cylinder 72 and the threaded spindle 70 have been drawn off because the broken-off part is identical to the working cylinder 32 in FIG. 2.

The axial shift brake bearing consists of two structurally identical support bearings 73 a and 73 b arranged symmetrically opposite each other in a cylinder tube 71 in a rotationally fixed and pressure-tight manner, which for the thread axial forces represent an axial slide bearing that is effective in two directions, the direction-changing thread axial forces with the axial disk 74 in the support bearings 73 a and 73 b for the threaded spindle 70 self-locking up to a braking torque specified by brake rings 75 a and 75 b, and the braking effect can be switched off, i.e. the threaded spindle 70 can then rotate as in axial roller bearings without any kind of inhibition.

Fig. 3 names in part because of the identical design and the symmetrical arrangement of certain components only once, in Fig. 3 these parts in addition to the identifying digits the letter "a" is added, z. B. support bearing 73 a. The complementary symmetrical part is identified by the same numerals in the following description, but the letter "b" is added. It should be noted that each a part named in FIG. 3 also includes an unnamed b part, which although mentioned in this description, does not have a corresponding reference symbol in FIG. 3.

The functional elements of a support bearing 73 a are arranged on a functional carrier 76 a. The spatial shape of the function carrier 76 a is that of a rotating body. The function carrier 76 a, which is provided with a centrally arranged, stepped, continuous central bore 77 a, consists of the support cylinder 78 a with the largest diameter and connected to it on both sides with flat stepped cylinder shoulders, whereby for reasons of a clear description, a smallest first cylinder heel is regarded as the start of the function carrier 76 a is that has the function of a ball bearing axis 79 a. This is followed axially by a second cylinder shoulder with the function of a ring plunger axis 80 a in relation to the ball bearing axis 79 a. This is followed axially by a third cylinder shoulder larger than the ring plunger axis 80 a with the function of a cylinder mount 81 a for a cylinder ring 82 a. This is followed axially by the larger and largest cylinder part of the function carrier 76 a with respect to the cylinder mount 81 a with the function of the support cylinder 78 a, which has the task of fixing the radial and axial position of the support bearing 73 a in the cylinder tube 71 . This is followed axially by a fourth cylinder shoulder, smaller in relation to the support cylinder 78 a, which, as the end cylinder 83 a of the function carrier 76 a, only has the task of reducing the installation length of the function carrier 76 a.

On the ball bearing axis 79 a is provided with a load surface 84 a rotating plate 85 a centrally rotatably mounted in a ball bearing 86 a with the load surface 84 a pointing to the beginning. The spatial shape of the rotating disk 85 a is that of a rotating body with a central bearing seat recess 87 a on the side facing away from the load area for the outer ring of the ball bearing 86 a, and a center arranged up to the load area 84 a continuous load area bore 88 a, into which the beginning of ball bearings axis 79 a hineinerstreckt and the load surface in the bore 88 a load face side, a seal ring groove 89 a and has arranged therein a sealing ring, wherein the load surface 84 a against the ball bearing shaft 79 a fluid-tight shut off. At the transition of the load surface 84 a to the outer diameter of the rotating disk 85 a, a chamfer serving as a braking surface 90 a is arranged. The inner ring of the ball bearing 86 a is fixed with a snap ring on the ball bearing axis 79 a, but finds its axial support on the ball bearing axis collar 91 a. The ring plunger axis 80 a is the inner radial pressure space boundary of a pressure medium-operated ring plunger 92 a guided thereon in a pressure-tight manner. The spatial shape of the ring plunger 92 a is that of a rotating body. The pressure medium is supplied through a collar bore 94 a in the axis collar 93 a of the ring plunger axis 80 a, the axis collar 93 a also being the axial limitation of the pressure space of the ring plunger 92 a. The collar bore 94 a is connected to a switching pressure bore 95 a arranged radially in the support cylinder 78 a, which leads further to a switching pressure connection 96 a arranged in the cylinder tube 71 . On the pressure medium-loaded plunger face 97 a of the ring plunger 92 a, sealing ring recesses are arranged in the inner and outer lateral surfaces with sealing rings 98 a and 99 a arranged therein, which seal the lateral surfaces. The side of the ring plunger 92 a opposite the pressurized plunger end 97 a has a cup-shaped recess 100 a with a guide bore 102 a arranged in the pot bottom 101 a, which seals and guides the ring plunger 92 a on the ring plunger axis 80 a, and thus also the inner lateral surface of the Represents ring plunger 92 a. The pot casing 103 a of the pot-shaped ring plunger 92 a extends without contact over the rotating disk 85 a up to the contact with a brake ring 75 a assigned to the braking surface 90 a of the rotating disk 85 a. The outer circumferential surface of the pot casing 103 a is guided and sealed in the cylinder ring 82 a arranged on the cylinder mount 81 a in a pressure-tight manner and on the collar 104 a.

The first and identical second support bearings 73 a and 73 b are symmetrically opposite one another, with the load surfaces 84 a and 84 b facing each other, and a distance between the load surfaces 84 a and 84 b, with their support cylinders 78 a and 78 b in one support cylinders 78 a and the same diameter 78b, the cylinder barrel diameter slightly enlarging Durchmesserausnehmung 105 of the cylinder tube 71 rotatably and fluid-tightly arranged, the collar surfaces 106 a and 106 b to the end cylinders 83 a and 83 b are one-sided truncated axially, once the end edge 107 of the Durchmesserausnehmung 105 and once by a retaining ring 108 , blown into a groove made in the diameter recess 105 . The cylinder rings 82 a and 82 b extend radially adjacent to the outer surface of the diameter recess 105 of the cylinder tube 71 up to a stop ring 109 arranged in the middle of the diameter recess 105 . Cylinder rings 82 a and 82 b, which act like precisely fitting spacer rings, in cooperation with stop ring 109 , support the opposite support cylinders 73 a and 73 b, which are supported on one side only, so that they are axially fixed on both sides.

The stop ring 109 arranged between the cylinder rings 82 a and 82 b is centered with its outer diameter on the outer surface of the cylinder tube 71 . The inner diameter is an outer guide ring surface for a multiplicity of compression springs 110 arranged in a circle in the circumferential direction and axially parallel, which act on brake rings 75 a and 75 b arranged on both sides of the stop ring 109 , the compression springs 110 being separated by intermediate, somewhat shorter (not shown) cylindrical rollers of the same diameter to prevent the compression springs 110 from getting caught. The spatial shape of the brake rings 75 a and 75 b is that of a rotating body, they are identical in construction and the brake rings 75 a and 75 b have on the spring-loaded flat ring surfaces 111 a and 111 b protruding inner guide edges 112 a and 112 b, which the location of the circularly arranged compression springs 110 and the cylindrical rollers arranged between them limit radially inwards.

The outer diameter of the brake rings 75 a and 75 b center on the cylinder rings 82 a and 82 b and are provided with two or more symmetrically opposite rectangular grooves 113 a and 113 b, which are compatible with radially equally wide, in the cylinder rings 82 a and 82 b are arranged slots 114 b and 114 b. In the chambers thus created, formed from the rectangular grooves 113 a and 113 b and the opposite slots 114 a and 114 b, circular disks 115 a and 115 b are arranged. The diameters of the discs 115 a and 115 b are selected so that they fill the chambers in the radial direction to a small degree of play, that is, the discs 115 a and 115 b act similarly to disc springs and a rotation of the brake rings 75 a and 75 b opposite prevent the cylindrical rings 82 a and 82 b. The thickness of the discs 115 a and 115 b is smaller by a small amount than the width of the slots 114 a and 114 b and rectangular grooves 113 a and 113 b, so that an axial displacement of the brake rings 75 a and 75 b is ensured. Compared to the ring surfaces 111 a and 111 b acted upon by compression springs 110 , the brake rings 75 a and 75 b have compatible friction surfaces 116 a and 116 b with respect to the braking surfaces 90 a and 90 b of the rotating disks 85 a and 85 b and these are caused by the the brake rings 75 a and 75 b acting axial spring forces on the braking surfaces 90 a and 90 b of the rotating disks 85 a and 85 b.

Between the load surfaces 84 a and 84 b, an axial disk 74 is arranged, which is fixedly connected to the thread-free shaft part 117 of the threaded spindle, which is mounted radially and pressure-tight in this through the central bore. The thread-free shaft part 117 extends through both function carriers 76 a and 76 b with one shaft end as the control shaft 118 emerging from the function carrier 76 a. The direction-changing thread axial forces are transmitted from the axial disk 74 to the corresponding load surfaces 84 a and 84 b of the rotating disks 85 a and 85 b. The axial washer 74 is designed such that central recesses 74 are arranged on both end faces of the axial washer, into which friction rings 119 a and 119 b are pressed in in a rotationally fixed manner above the end face, so that the friction rings 119 a and 119 b ring end faces on the axial washer 74 form and with axial load only the ring end faces with the load surface 84 a or 84 b contact. With an axial load, the axial disk 74 provided with friction rings 119 a and 119 b forms a so-called axial pressure chamber 120 a or 120 b with the load surfaces 84 a or 84 b. The axial pressure chambers 120 a or 120 b can be pressurized through the central bores 77 a and 77 b in the function carriers 76 a and 76 b. For this purpose, part of the central bores 77 a and 77 b used to support and seal the shaft part 117 is somewhat enlarged from the beginning of the ball bearing axis 79 a and 79 b, so that annular spaces 121 between the shaft part 117 and the central bores 77 a and 77 b a and 121 b arise as a flow channel for the pressure medium. The annular spaces 121 a and 121 b extend in the function carriers 76 a and 76 b up to a connection to the control pressure bores 122 a and 122 b arranged radially outwards in the support cylinders 78 a and 78 b, which further to the ones arranged in the cylinder tube 71 Guide control pressure connections 123 a and 123 b in cylinder tube 71 .

As mentioned at the beginning of the description of FIG. 3, an axial brake bearing is largely similar to the described axial shift brake bearing. In the case of the axial brake bearing, the brake rings 75 a and 75 b are in permanent brake contact with the braking surfaces 90 a and 90 b of the rotating disks 85 a and 85 b. Switching off the braking is not possible, that is to say that the axial brake bearing has no ring plungers 92 a and 92 b, and the bores in the support cylinders 78 a and 78 b necessary for acting on the ring plungers 92 a and 92 b are also not present. A constructively possible shortening of the overall length of the axial brake bearing compared to the axial shift brake bearing is irrelevant for the structural design, and a further description is therefore unnecessary.

The function of the devices according to the invention according to FIG. 3 is explained in more detail below.

It is assumed that a spindle nut, not shown, arranged on the non-self-locking threaded spindle is connected to a working piston and transmits the piston force in direction 124 to the threaded spindle 70 , into which a spindle torque is thereby introduced, but which does not cause rotation because at the same time the friction ring 119 a of the axial disk 74 is pressed onto the load surface 84 a of the rotating disk 85 a and here it is assumed that the braking torque acting on the rotating disk 85 a of the brake ring 75 a is greater than that introduced into the threaded spindle, that is, as the inhibiting torque transferred to the load surface 84 a. As long as the rotating disk 85 a is held by the brake ring 75 a, the function between the axial disk 74 and the load surface 84 a is identical to that of FIGS . 1 and 2 in relation to the design of the self-locking, the mode of operation of the axial pressure chambers 19 a and 58 a and the control shaft 15 and 55 described. By appropriate pressurization of the axial pressure chamber 120 a, the self-locking can also largely be switched off in the embodiment shown in FIG. 3. However, if the relief pressure in the axial pressure chamber 120 a exceeds the supporting force on the load surface 84 a, the axial disk 74 lifts off from the load surface 84 a, with the result of a considerable pressure loss due to pressure medium flowing out, and pneumatic loading also creates self-excited pneumatic instabilities that are known under the term "air hammer". A complete deactivation of any kind of inhibition without the disadvantages described above is possible if the ring plunger 92 a is acted upon by pressure medium and the brake ring 75 a is pressed off from the braking surface 90 a of the rotating disk 85 a. The ball-bearing rotating plate 85 a then causes an axial fixation of the threaded spindle, as if it were itself low-friction in bearings. If only one Ringplunger 92 applied to a, the result is for the threaded spindle an only one-sided off the inhibition, or only a unidirectional self-locking, that is, the working piston connected to the spindle nut can be moved by the pressure medium freely in one direction. Movement in the opposite direction is prevented by self-locking.

Reference list

Fig. 1
1 working cylinder
2 control gears
3 working pistons
4 piston rod
5 output connection
6 socket
7 connecting bridge
8 frame
9 threaded spindle
10 spindle nut
11 thrust bearing
12 frame part
13 load area a and b
14 friction rings a and b
15 control shaft
16 abutments a and b
17 axial washer
18 shaft part
19 axial pressure chamber a and b
20 control pressure connection a and b
21 work space
22 Pressure fluid connection
23 direction

Fig. 2
30 cylinder covers
31 thrust bearing
32 working cylinders
33 working pistons
34 piston rod
35 threaded spindle
36 spindle nut
37 abutments a and b
38 thrust washer
39 work space
40 shaft part
41 central bore a and b
42 support cylinders a and b
43 cylinder paragraph a and b
44 plane surface a and b
45 load area a and b
46 annular space a and b
47 radial bore a and b
48 Control pressure connection a and b
49 spacer ring
50 diameter recess
51 cylinder tube
52 end edge
53 circlip
54 cavity
55 control shaft
56 end faces a and b
57 Friction rings a and b
58 axial pressure chamber a and b
59 Pressure fluid connection
60 direction

Fig. 3
Part arranged in parentheses (b) symmetrically, which has no reference numerals in FIG. 3, but is mentioned in the description
70 threaded spindle
71 cylinder tube
72 working cylinders
73 support bearings a and (b)
74 thrust washer
75 brake rings a and (b)
76 function carriers a and (b)
77 central bore a and (b)
78 support cylinders a and (b)
79 ball bearing axis a and (b)
80 ring plunger axis a and (b)
81 cylinder socket a and (b)
82 cylindrical ring a and (b)
83 end cylinders a and (b)
84 load area a and (b)
85 turntable a and (b)
86 ball bearings a and (b)
87 bearing seat recess a and (b)
88 load area bore a and (b)
89 sealing ring groove a and (b)
90 braking surface a and (b)
91 ball bearing axle collar a and (b)
92 ring plungers a and (b)
93 axis bundle a and (b)
94 collar bore a and (b)
95 switching pressure bores a and (b)
96 switching pressure connection a and (b)
97 plunger end face a and (b)
98 sealing ring 98 a and (b)
99 sealing ring 99 a and (b)
100 recesses a and (b)
101 pot base a and (b)
102 guide bore a and (b)
103 pot jacket a and (b)
104 framing a and (b)
105 diameter recess
106 fret areas a and (b)
107 end edge
108 circlip
109 stop ring
110 compression springs
111 ring surfaces a and (b)
112 leading edges a and (b)
113 rectangular grooves a and (b)
114 slots a and (b)
115 discs a and (b)
116 friction surfaces a and (b)
117 shaft part
118 control shaft
119 friction rings a and (b)
120 axial pressure chamber a and (b)
121 annulus a and (b)
122 pilot pressure bores a and (b)
123 control pressure connection a and (b)
124 direction

Claims (20)

1. Pressure-operated cylinder with the following features:

• a) a working piston ( 3 ) is arranged axially displaceably in the working space ( 21 ) of the working cylinder ( 1 ),
• b) the working piston ( 3 ) can be acted upon on one side and / or on both sides with pressure medium, whereby the working piston ( 3 ), which can only be acted upon on one side, can also have the shape of a plunger and the working cylinder ( 1 ) is closed with only one cylinder cover,
• c) the working piston ( 3 ) is positively or non-positively connected to coupling elements, which utilize and / or tap the movements of the working piston ( 3 ) out of the working space ( 21 ), the coupling elements, for example, piston rods ( 4 ) guided out of the cylinder covers, Belts, ropes or slides arranged axially parallel to the working piston ( 3 ), the slides producing a coupling connection to the working piston ( 3 ) by means of controlled slots in the jacket of the working cylinder in a form-fitting manner with webs or non-positively with magnetic forces,
• d) the work area cross section is circular, rectangular, square, oval or has any suitable cross section,

characterized by the following features:

• e) the working piston ( 3 ) is connected to a spindle nut ( 10 ) screwing on a threaded spindle ( 9 ), the spindle nut ( 10 ) being rotationally fixed with respect to the working cylinder ( 1 ), but axially displaceable by the working piston ( 3 ) and the threaded spindle ( 9 ) is stationary but rotatable with respect to the working cylinder ( 1 ),
• f) the direction of movement transmitted from the working piston ( 3 ) to the spindle nut ( 10 ) is axially parallel or axially identical to the threaded spindle ( 9 ),
• g) the threaded spindle ( 9 ) is self-locking with respect to an axial force introduced into the threaded spindle ( 9 ) by the spindle nut ( 10 ),
• h) the threaded spindle ( 9 ) is assigned a control drive driving or regulating the threaded spindle ( 9 ),
• i) the control drive is motorized and / or powered by muscle power.

2. Pressure medium operated working cylinder ( 32 ) with the following features:

• a) the working cylinder ( 32 ) is closed on both sides with a first ( 30 ) and a second cylinder cover,
• b) a working piston ( 33 ) is arranged axially displaceably in the working space ( 39 ) of the working cylinder ( 32 ),
• c) the working piston ( 33 ) can be pressurized on both sides,
• d) the working piston ( 33 ) is connected to a hollow cylindrical piston rod ( 34 ) which is closed on one side and which is guided out of the first cylinder cover ( 30 ) with the closed side,
• e) a threaded spindle ( 35 ), which is mounted rotatably and is fixed in the axial direction with respect to the working cylinder ( 32 ), extends into the piston rod ( 34 ),
• f) the threaded spindle ( 35 ) is rotatably mounted and axially fixed in the second cylinder cover,
• g) a spindle nut ( 36 ), which screws with the threaded spindle ( 35 ), is arranged in a rotationally fixed manner in the working piston ( 33 ),
• h) the working piston ( 33 ) is arranged in a rotationally fixed manner with respect to the working cylinder ( 32 ),
• i) the cross section of the work area is circular, rectangular, square, oval or has any suitable cross section,

characterized by the following features:

• j) the threaded spindle ( 35 ) is self-locking with respect to an axial force introduced into the threaded spindle ( 35 ) by the spindle nut ( 36 ),
• k) the threaded spindle ( 35 ) is assigned a control drive driving or regulating the threaded spindle ( 35 ),
• l) the control drive is motorized and / or powered by muscle power.

3. Device according to claim 1 and 2 characterized in that the self-locking of the threaded spindle (9, 35, 70) caused when a non-self-locking threaded spindle (9, 35, 70) a Axialhemmlager (11 and 31) is assigned. 4. Device according to claim 3, characterized in that the axial locking bearing ( 11 and 31 ) is controllable and / or adjustable within the limits of self-locking to non-self-locking. 5. Device according to claim 3, characterized in that the thrust bearing ( 11 and 31 ) can be switched uninhibited. 6. Device according to claim 3, characterized in that the thrust bearing ( 11 and 31 ) is designed so that between two spaced abutments ( 16 a and 37 a) with mutually facing opposite load surfaces ( 13 a, 45th a and 84 a) an axial disc ( 17 , 38 and 74 ) is arranged, which is firmly connected to the threaded spindle ( 9 , 35 and 70 ) and changes the axial axial thread forces on both sides of the front face of the axial disc ( 17 , 38 and 74 ) onto the load surfaces ( 13 a, 45 a and 84 a) of the abutment ( 16 a and 37 a) transmits and represents a two-directional axial plain bearing for the threaded axial forces, the effective friction diameter of the end faces by the choice of a correspondingly large diameter of the axial disk ( 17 , 38 and 74 ) is selected so large that the resulting frictional torque from the frictional torque between the spindle nut ( 10 and 36 ) and the threaded spindle ( 9 , 35 and 70 ) and the inhibiting torque on the load surfaces ( 13 a, 45 a and 84 a) is self-locking for the threaded spindle ( 9 , 35 and 70 ). 7. Device according to claim 6, characterized by the following features:

• a) on the end faces on both sides of the axial disk ( 17 , 38 and 74 ) recesses are arranged centrally, which each extend from the center of the end faces to almost the outer edge of the end faces, so that only narrow ring end faces rubbing on the flat load surfaces ( 13 a, 45 a and 84 a) the abutment ( 16 a and 37 a) can act, and thus the effective friction diameter is pretty much in the middle of the ring face,
• b) when the annular end face is supported on the load faces ( 13 a, 45 a and 84 a), the axial pressure chambers ( 19 a, 58 a and 120 a), which are arranged centrally in the end faces, form pressurized axial pressure chambers ( 19 a, 58 a and 120 a), the axial pressure chambers ( 19 a, 58 a and 120 a) are an element for controlling or regulating the friction between the ring end faces and the load surfaces ( 13 a, 45 a and 84 a),
• c) for loading the axial pressure chamber ( 19 a, 58 a and 120 a) with pressure medium, holes are arranged in the load areas ( 13 a, 45 a and 84 a) with control pressure connections ( 20 a, 48 a and 123 a).

8. Device according to claim 7, characterized in that from the sealing action of the ring end faces when rubbing on the load surfaces ( 13 a, 45 a and 84 a) by slightly flowing pressure medium lubrication and cooling of the load surfaces ( 13 a, 45 a and 84 a ) arises. 9. Device according to claim 1 to 7, characterized by the

• a) the control drive or the control shaft ( 15 , 55 and 118 ) is assigned a valve arrangement and a device, the device triggering a circuit for switching on or initiating a control torque for generating a control rotation simultaneously for the valve arrangement which continuously fills the axial pressure chamber with pressure medium ,
• b) the control drive or the control shaft ( 15 , 55 and 118 ) is associated with a valve arrangement and a device, the device, after switching on or initiating a control torque for generating a control rotation at the start of a control rotation, simultaneously triggering a circuit for the valve arrangement which continuously pressure medium discharges from the axial pressure chamber up to a previously set pressure.

10. Device according to claim 1 to 7, characterized in that the control drive or the control shaft ( 15 , 55 , 118 ) of the threaded spindle ( 9 , 35 and 70 ) is assigned at least one torque-measuring device and the direction of the torque detecting device and this device Evaluation circuit is assigned, which generates representative analog and / or digital control signals depending on the detected values. 11. The device according to claim 10, characterized in that the Evaluation circuit with at least one valve for control and / or Rules of the pressure medium is connected.   12. The device according to claim 1 to 7, characterized in that the control drive or the control shaft ( 15 , 55 , 118 ) of the threaded spindle ( 9 , 35 and 70 ) at least one, the pressure medium depending on the torque and direction of rotation, and / or controlling device is assigned. 13. The device according to claim 6 to 8, characterized in that the load surfaces ( 84 a) are arranged on support bearings ( 73 a) rotatably mounted rotary disks ( 85 a), each of which is assigned a torque-limiting brake. 14. Device according to claim 13, characterized in that on the load surfaces ( 84 a) of the rotating disks ( 85 a) annular or conical annular braking surfaces ( 90 a) are arranged centrally and the braking surfaces ( 90 a) are rotationally fixed by spring force on the braking surfaces ( 90 a) pressing brake rings ( 75 a) that are compatible with the braking surfaces are assigned as torque limiters. 15. The device according to claim 13 to 14, characterized in that the brake pressing forces act on the braking surfaces ( 90 a) of the rotating discs ( 85 a) in the same direction as the axial force of the axial disc ( 74 ) on the load surface ( 84 a) the turntable ( 85 a). 16. The device according to claim 13 to 15, characterized in that the rotary disks ( 85 a) are associated with devices for brake ventilation. 17. The device according to claim 13 and 16, characterized in that the rotating disks ( 85 a) pressure medium-operated devices for brake ventilation are assigned. 18. Device according to claim 13 and 18, characterized in that the brake rings ( 75 a) with ring plunger axles ( 80 a) guided and pressurized with ring plunger ( 92 a) for brake release are assigned.