DE19633420A1 - Clutch control system particularly for motor vehicle - Google Patents

Abstract

The clutch is operated by a hydraulic slave cylinder (6) with the hydraulic pressure regulated by a proportional control valve (2). The servo pressure is applied in a programmed characteristic over the movement of the clutch to provide a controlled variation of the torque. The system operates automatically or manually. The control takes into account the temperature of the hydraulic fluid and the state of the engine e.g. it makes allowances for cold starting. The transmission slip between the drive and the driven sections of the system is monitored to provide the optimum torque change. The proportional control valve is controlled by a solenoid control.

Description

The invention relates to a method for hydraulic Actuation of a clutch according to the preamble of the patent say 1 and to a device for performing this Method according to the preamble of claim 10 or 12. In particular, the invention relates to a method and a device for automatic hydraulic actuation supply of a motor vehicle clutch.

Automatic transmissions for motor vehicles with combustion engines have come in various designs in the USA and Japan prevailed, where the market share for such Ge drives are now between 75% and 85%. There are many Reasons in favor of the trend in Europe too higher automation of motor vehicles to be expected is. In particular, this is due to the increasing traffic density higher switching frequency due to actuation of the clutch the clutch pedal and shifting the transmission via the shift lever more and more to the chore. In addition, in terms of to the increasingly strict emissions and noise regulations the internal combustion engine targeted in favorable operation points to be operated, for example in the warm running phase to achieve the prescribed values, whereby the on Motor vehicle partly with considerable device technology Measures taken to comply with the prescribed values, e.g. B. the purification of exhaust gases from combustion engine with the help of a regulated catalyst, conditional through the free choice of the gear ratio by the Driver of the motor vehicle partially not or not effectively can be implemented.  

Automatic clutch actuation to disengage a friction mechanism Lagkupplung on a motor vehicle are in the prior art already known and will be installed in series.

For example, DE 37 06 849 A1 describes DE 37 06 850 A1 or DE 41 20 128 A1 automatic hydraulic Clutch actuation, in which one with the release bearing of Coupling clutch slave cylinder of one Clutch master cylinder is hydraulically controlled to the to apply the force required to disengage the clutch. According to this state of the art is the clutch master cylinder with provided with a mechanical actuator that has a Worm gear driven by an electric motor to actuate the clutch master cylinder. As a result, the In principle, clutch independent of a clutch pedal auto be matically disengaged.

However, this prior art has the disadvantage that a considerable amount of mechanical effort a relatively large amount of space must be operated in order to look at the wear of the relatively moving components correct operation of the clutch actuation permanently guarantee.

Furthermore, from DE 36 30 750 A1 or DE 41 21 016 A1 automatic clutch actuation known in which at the Kupp a pedal is attached to an electro African control unit is connected. In the case of DE 36 30 750 A1 controls the control unit depending on the signal of the encoder a switch valve electrically to one with the Clutch release bearing operatively connected piston-cylinder connection order optionally to connect with a hydraulic pump, see above that this by pressurizing the piston-cylinder assembly force required to disengage the clutch. In the case of DE 41 21 016 A1, the control unit controls in Ab depending on the signal from the position sensor a volume-proportional control valve electrically to a  piston-cylinder connected to the clutch release bearing the arrangement via the valve assembly with one of a Hy to connect the hydraulic pump fed pressure accumulator so that the force required to disengage the clutch Pressurization of the piston-cylinder arrangement applied becomes.

Problems exist with the prior art described above in that a relatively large device technology Effort with fast responding valves and accordingly precise displacement measuring device on the piston-cylinder arrangement must be operated to prevent the piston cylinder the arrangement and thus the clutch oversteered impermissibly will. A fine dosage of the release force is also necessary Avoidance of coupling shocks, for example, is not possible or at least difficult.

The same problems also occur with those from DE 42 37 853 A1 and DE 43 09 901 A1 known automatic clutch actuators activities. According to this prior art, one with the Clutch release cylinder, piston-cylinder arrangement, which is operatively connected Hydraulically controlled by an actuating cylinder to the to apply the force required to disengage the clutch. The The cylinder chamber of the actuating cylinder is in by a piston divided a servo room and a work room, one of which the latter hydraulically ver with the piston-cylinder arrangement is bound. The pressure in the work area and thus in the piston Cylinder arrangement is determined by the pressure in the servo room, which in turn has a volume-proportional control valve is controllable. An electronic control system speaks this is the current position of a piston rod of the Stellzylin derkolbens and thus the current position of the release of the Coupling signal representing a displacement sensor on and control is proportional to the desired volume via the control valve Position of the release valve the pressure in the servo space of the Stellzylin otherwise.  

After all, automatic switching operations are for power vehicles, with the help of a gear shift stage can be switched on or switched off automatically while standing the technology known and are also installed in series. For this include automatic transmissions with hydraulically controlled tarpaulin gear sets and hydrodynamic converter, continuously variable transmission with and without switching stages and starting clutch as well as double clutch power transmission. While in automatic transmissions with hydraulic controlled planetary gear sets and hydrodynamic converter even after long optimization and the use of complex walls bridging clutches the achieved efficiency of the Mo ment transfer still significantly lower than before directional effort is considerably higher than by hand shift-driven, the efficiencies of the moments are transmission with stepless transmissions even lower than with conventional automatic transmissions. Double clutch gearboxes, however, are still essential in terms of device technology more complex than conventional automatic transmission, whereby be correct switching sequences cannot be carried out directly.

In summary, it should be noted that the known automa table clutch or shift operations compared to the conventional manual operations with regard to their type speaking and operating behavior taking into account one measured device engineering effort, especially ge low costs and small installation space, still improvement are in need.

Compared to the prior art described above, the The invention is therefore based on the object, based on the prior art Technology according to, for example, DE 43 09 901 A1, a Ver drive and a device for automatic actuation of a To create coupling with low device technology Effort an improved response and operating behavior of the Enable clutch.  

This object is achieved by the in claim 1 and in Claim 10 or 12 specified Merkmaie solved. Before partial and / or expedient developments of the invention are the subject of claims 2 to 9, 11 and 13 to 18.

According to the hydraulic actuation of the the clutch slave cylinder operatively connected clutch in which a servo pressure is generated in the servo circuit which the hydraulic control of the clutch master cylinder serves to generate a working pressure in the pressure circuit, which is applied to the clutch slave cylinder, or is placed directly on the clutch slave cylinder, to the clutch slave cylinder to disengage the clutch required power to apply, the servo pressure to turn on and off back of the clutch set in the servo circuit. To is the servo circuit with a proportional valve according to the invention til provided, by means of which the servo pressure in the servo circuit de finely adjustable.

Through the defined setting or control of the servo it is in the servo circuit - in contrast to the state of the art nik, according to which the volume flow of the pressure medium in the servo circuit controlled by a volume-proportional control valve becomes - possible, the one required to disengage the clutch Apply force in a gentle manner to the clutch, so that a good response and operating behavior of the clutch, like a optimally on the motor vehicle or its operating state tuned engagement process of the clutch, a smooth ran yaw mode with sliding clutch, "cavalier start" with sliding clutch without unduly excessive stress the coupling, etc., can be guaranteed. Especially have volume fluctuations in the pressure medium in the servo or Pressure circuit caused by temperature influences, compressibility losses due to air and water content of the pressure medium, Fluid loss through the compensating hole in the clutch master cylinder linders, "tongue flap" on the pad springs and / or axial  impact of the crankshaft may have no influence the control of the engagement and disengagement process.

In addition, the conventional with the clutch master cylinder Licher rigidly connected clutch pedal can be omitted the pedal modules are advantageously mounted in a simplified manner be, whereby only a small bulkhead cutout not manoeuvrable and the connection of the hydraulic lines in pre-assembly is possible. Furthermore, by simple exchange or Ent remove the clutch master cylinder and provide the servo circuit tried and tested hydraulic clutch actuation will. Finally, compared to the conventional ones hydraulic clutch actuation great space advantages, in particular greater flexibility, since the components of the Servo circuit decentralized from the clutch master cylinder or clutch mer cylinder can be arranged.

According to claim 2, when disengaging a clutch, whose characteristic of the disengagement force over the disengagement path in essentially like a parabola open at the bottom (so-called "drop-off" characteristic), the release path or a size proportional to this together with a slip representing drive and output of the clutch Size captured and clutch engagement based on the Regulated sizes regulated so that a gentle indentation of the Coupling is possible. The location of the Piston in the clutch master cylinder or the position of the piston in the Clutch slave cylinder than the clutch release path proportional size, as indicated in claim 3.

According to claim 4, when disengaging a clutch, whose characteristic of the disengagement force over the disengagement path with increasing release path, the release force or a size proportional to this together with a slip representing drive and output of the clutch Size captured and clutch engagement based on the Regulated sizes regulated. This regulation also enables  the clutch engages gently. In addition, at this regulation is a travel sensor for recording the release path or a proportional size and an elaborate after position advantageous when the clutch friction linings wear out omitted. With such a force proportional control of the engagement process arise compared to a path-dependent Regulation further advantages in that disturbances, such as Thermal expansion of the clutch pressure plate, the friction disc, the Diaphragm spring, etc. or by axial vibrations of the crank shaft caused superimposed movements on the coupling or Clutch actuation, have no influence on the control.

Couplings with a steadily increasing characteristic of the Disengagement force can be built up more easily than clutches whose Release force characteristic has a "drop-off" characteristic.

According to claim 5 for such a scheme Working pressure in the pressure circuit or the servo pressure in the servo circuit than the size proportional to the clutch release force be applied. However, the coil current preferably serves one Magnetic drive one in the servo circuit to control the servo pressure provided proportional valve or the coil current one for driving a hydraulic pump in the servo circuit seen electric motor than the clutch release force proportional size, as indicated in claim 6, because no additional sensors are required to record these quantities be done.

According to claim 7, the engagement of a clutch, whose characteristic of the disengagement force over the disengagement path in the we considerably how a parabola opens downwards, in discontinuous part of the characteristic curve according to the method claim 2 or 3 and in the continuous sub-area of Characteristic curve advantageous according to the method according to one of the Pa claims 4 to 6 regulated. With such a mixed procedure Ren is thus the engagement of the clutch as long as Basis of the detected disengagement path or the proportionate to it len size regulated until the course of the characteristic over the  Release path is clear. Then the control is switched and based on the detected disengagement force or proportional size continued to be a special to ensure sensitive engagement of the clutch.

The claim 8 provides for the application of the dome to coordinate such that a torque tracking ge will create, according to which the engine torque to the Coupling connected drive determined and dependent a servo based on the determined motor torque in the servo circuit pressure, which is set directly or via the clutch the master cylinder by means of the clutch slave cylinder on the Clutch applies a force that transmits a sol Chen torque made possible by the clutch, the smaller is the maximum torque that can be transmitted by the clutch, but is larger than that by a predetermined amount determined engine torque. As a result, the for the disengaging process required speed of the Clutch pressure plate, so that even with a fast Gear changes can be completely disengaged in time.

According to claim 9, the servo circuit is dependent speed of an initially determined engine torque did not set servo printing until a later determined engine torque by a predetermined amount of the initially determined engine torque deviates, so that the Control of the servo pressure in the servo circuit is simplified.

The hydraulic effective surface of the piston of the Clutch master cylinder can, depending on the circumstances of the ar the hydraulic effective area of the piston on the working space side is the same or may also be smaller than this, however, the Arrangement taken according to claim 11, in which the water hydraulic effective surface of the clutch piston on the fore-end master cylinder larger than the workspace-side hydraulic lische effective surface of the piston is because of this on simple  A hydraulic ratio or reinforcement in the dome donor cylinder is realized.

According to claim 13, the proportional valve in the Ser vokreis a valve body that one when the clutch with a pressure medium through which the valve gap flows limited electromagnetically controllable as the servo pressure Back pressure in a pressure chamber of the proportional valve defi set with the servo space of the clutch encoder cylinder or the clutch slave cylinder is connected.

According to claim 14, the pressure chamber of the Proportio nalventils connected to a hydraulic pump, the at Betä the pressure medium through the valve gap of the clutch Proportional valve promotes. Accordingly, it can be more advantageous A hydraulic pump already present in the motor vehicle principally serve to generate the servo pressure.

The claim 15 provides that the servo circuit in Locked zero position biased proportional valve that one with a pressure source and the servo space of the clutch master cylinder or the clutch slave cylinder hydraulically connected pressure chamber and a valve body. Of the Valve body limits you when engaging the clutch a pressure gap through which the valve gap is electroma electroma gnetisch controllable, so that the servo pressure as a dynamic pressure in the pressure chamber can be set and is defined against the pressure in the pressure chamber by means of a closing spring biased, the biasing force is equal to the force that on Valve body acts on the pressure chamber side when disengaging the coupling requires the maximum working pressure in the pressure circuit or servo printing in the servo circuit. As a result, he needs Magnetic drive in the proportional valve when disengaging the clutch tion are not supplied with current, since the generator used to generate the Disengaging the clutch maximum required working pressure required servo pressure or the maxi to disengage the clutch  times required servo pressure against the force of the closing spring can build up in the pressure chamber of the proportional valve.

According to claim 16, the proportional valve is in front preferably as a ball seat valve with a spherical valve body formed together with an annular Ven tilsitz limits the valve gap of the proportional valve. Like this with the servo printing can be done precisely via an inexpensive and small proportional valve can be set, its location not to a specific point in the motor vehicle, for example at the hitch master cylinder or on the clutch slave cylinder, bound, but is freely selectable.

Claim 17 provides one to the servo circuit Connect the circuit, the hydraulic actuation a manual gearbox, what the circuit for each Switching stage of the gearbox a shift cylinder and a Proportio connected to a pressure side of the servo circuit has valve for controlling the shift cylinder, and has a switching valve, the circuit with either a Suction side of the servo circuit connects. According to the claim 18, the proportional valve of the circuit is preferred as an electromagnetically controllable ball seat valve that is biased in the locked zero position.

It is thus advantageously possible to use the in the servo circuit generated and defined servo pressure for hydrau actuation of proven mechanical manual transmission with good efficiency of torque transmission. Here any switching sequences can be implemented without problems. Also the manual transmission is protected because the change of gear stage can be done in a defined manner, with an individual Adjustment of the actuation forces for the manual transmission the proportional valves in the hydraulic circuit are possible is. Finally, the conventional shift linkage can ent fall so that the assembly of the drive train is facilitated and space advantages result.

The invention is described below based on a preferred embodiment examples explained in more detail with reference to the drawing, the same or similar parts with the same reference numerals are provided. Show:

Fig. 1 shows a schematic representation of a first embodiment of the hydraulic Kupp inventive lung activity with a clutch master cylinder between Ser and vokreis pressure circuit, as well as a proportional valve and a hydraulic pump for generating and continuous adjustment of the servo pressure in the servo circuit

Fig. 2A is a coupling characteristic with a so-called "drop-off" characteristic in which the clutch torque M, the release load F and the coil current I _V of a magnetic drive in the proportional valve of the servo loop or the current I _M of an electric motor for driving the hydraulic pump in the servo loop as a function the disengagement path a of the clutch pressure plate are worn,

Fig. 2B shows a of FIG. 2A corresponding representation of a coupling characteristic with a substantially linear or constantly increasing characteristic,

Fig. 3 shows a schematic representation of a variant of, shown in FIG. First embodiment of the hydraulic clutch control system according to the invention, in which a different proportional valve in the servo circuit for use ge reached 1

Fig. 4 shows a schematic representation of a second embodiment of the hydraulic Kupp inventive lung activity without clutch master cylinder, in which the Ser is vokreis connected directly to the clutch slave cylinder, but otherwise the servo circuit of FIG. 1 corresponds,

Fig. 5 shows a schematic representation lung activity of a third embodiment of the hydraulic Kupp invention, the activity compared to the first embodiment of FIG. 1 with an additional hydraulic circuit for shifting a mechanical transmission having a plurality of gear stages is provided,

Fig. 6 is a schematic representation of the provided in each additional hydraulic circuit for switching operation according to FIG. 5 for each switching stage proportional valve in an enlarged representation, and

Fig. 7 is a schematic representation of a connection between the supply in the additional hydraulic circuit for Schaltbetäti supply according to FIG. 5 provided for each switching stage of the mechanical gear shift cylinder and a sliding sleeve of the mechanical transmission.

Fig. 1 shows the first embodiment of the inventive hydraulic clutch actuation in the unactuated state. Referring to FIG. 1, the hydraulic clutch actuation system has a clutch master cylinder 2, which is connected via a pressure line 4 with a clutch slave cylinder 6 to form a pressure circuit D, so that is operatively connected to the clutch slave cylinder 6 clutch pressure plate (not shown) by operating the Kupp lung master cylinder 2 with a Release force can be applied to lift the clutch pressure plate from a clutch disc (not shown) and thus disengage the clutch.

Furthermore, the hydraulic clutch actuation has a hy draulic servo circuit S for controlling the pressure circuit D via the clutch master cylinder 2 . The hydraulic servo circuit S comprises the clutch master cylinder 2 , a proportional valve 8 as a pressure generator unit, a hydraulic pump 10 which can be driven by an electric motor 12 or an electromagnetic clutch, and a reservoir 14 for the pressure medium, such as brake fluid, mineral oil, etc. The interconnection of these components will be discussed in more detail after the following description of the individual components.

The clutch master cylinder 2 has a housing 16 with a pressure compensation and follow-up device 18 , which is formed on a Ge housing bore 20 of the housing 16 , in which a hydraulically sealed piston 22 is axially displaceable and is guided on the housing bore 20 . The piston 22 is provided at its end portions on the outer circumference, each with a sealing element 24.1 , 24.2 and tapered between the Dichtele elements 24.1 , 24.2 . The piston 22 divides the housing bore 20 of the clutch master cylinder 2 into a working space 26 which is delimited by the one end face of the piston 22 , the left sealing element 24.1 in FIG. 1 and the housing 16 , and a trailing space 28 which concentrically surrounds the piston 22 and extends radially outward is delimited by the housing 16 and on both sides by the sealing elements 24.1 , 24.2 , and a servo or amplifier space 30 which is fluid-tight from the other end face of the piston 22 , the sealing element 24.2 on the right in FIG. 1, and the housing 16 attached housing base 32 is limited. In the working space 26 a compression spring 34 is arranged, which holds the piston 22 in the unactuated state, ie in its basic position against the stop on the housing base 32 . In this basic position of the piston 22 , the pressure compensation and trailing device 18 connects the trailing space 28 with the working space 26 of the clutch master cylinder 2 .

Further, the clutch master cylinder 2 depending on builds character Route disengaging-course ristik the clutch pressure plate, is discussed in greater detail the, a displacement sensor 36 arranged by means of which the position of the piston 22 in the housing bore 20 of the clutch master cylinder 2 can be determined. The displacement sensor 36 can, however, just as well be arranged in the clutch master cylinder 2 or on or in the clutch slave cylinder 6 .

The serving as a pressure generator proportional valve 8 is formed as an electromagnetically actuated 2/2 ball seat valve, which is switched to the zero position in the uncontrolled state. The proportional valve 8 has a three-chamber valve housing 38 , which has an armature chamber 40 , an outlet chamber 42 and a pressure chamber 44 , which are arranged in this order in Fig. 1 from left to right behind each other concentrically to the central axis of the valve housing 38 .

In the armature chamber 40 of the proportional valve 8 , a control piston 46 is received, which consists of a ferromagnetic armature 48 and a valve bolt 50 mounted centrally on this. In the radial wall of the armature chamber 40 , a solenoid 52 is provided, which surrounds the armature 48 at least partially concentrically and together with the armature 48 forms a magnetic drive with the aid of which the control piston 46 can be displaced in a defined manner in the axial direction of the armature chamber 40 . The valve pin 50 of the control piston 46 is hydraulically sealed in a partition 54 of the valve housing 38 between the armature chamber 40 and the drain chamber 42 and projects into the drain chamber 42 .

In the flow chamber 42, a valve body 56 is arranged which is formed as a metallic sphere therethrough over which abge sealed by the partition wall 54 between the armature chamber 40 and the discharge chamber 42 valve pin 50 of the control piston 46 can be mechanically connected to an actuating force beauf beat. The valve body 56 can be struck on the valve bolt 50 of the control piston 46 with a compressive force only, since the valve body 56 and the valve bolt 50 are two separate components.

The pressure chamber 44 , which is smaller in diameter, adjoins the drain chamber 42 . At the discharge chamber end of the pressure chamber 44 , an annular valve seat 58 is formed centrally to the central axis of the control piston 46 , which together with the valve body 56 defines a valve gap 60 , the flow cross section of which corresponds to the throttle cross section of the proportional valve 8 . In the pressure chamber 44 a return spring 62 is arranged, which presses the valve body 56 against the valve pin 50 of the control piston 46 .

In the zero passage position of the proportional valve 8 , the valve gap 60 between the valve body 56 and the valve seat 58 is maximally opened by the restoring force of the return spring 62 , the control piston 46 indirectly via the valve body 56 against a stop 64 facing away from the drain chamber 42 End of the armature chamber 40 is pressed. The maximum possible stroke of the armature 48 of the control piston 46 in the armature chamber 40 corresponds at least to the closing travel of the valve gap 60 , so that by axially displacing the control piston 46 via the best consisting of armature 48 and solenoid 52 existing magnetic drive of the valve gap 60 can be set can.

The clutch master cylinder 2 and the proportional valve 8 are connected to the hydraulic pump 10 and the reservoir 14 as follows.

The booster chamber 30 of the clutch master cylinder 2 is connected via a pressure line 66 to the pressure chamber 44 of the proportional valve 8 , which is connected via a pressure line 68 to the output of the hydraulic pump 10 . The input of the hydraulic pump 10 is connected via a suction line 70 to the reservoir 14 before. The reservoir 14 is connected via a drain line 72 to the drain chamber 42 of the proportional valve 8 , which is connected via a follow-up line 74 to the follow-up chamber 28 of the clutch master cylinder 2 . For the person skilled in the art it can be seen that the pressure line 66 does not have to be connected to the pressure line 68 via the pressure chamber 44 of the proportional valve 8 , but can also branch off directly from the pressure line 68 . Likewise, the wake line 74 need not be connected to the drain line 72 via the drain chamber 42 of the proportional valve 8 , but can branch off from the drain line 72 or be connected directly to the reservoir 14 .

Finally, the displacement sensor 36 of the clutch master cylinder 2 , the solenoid 52 of the proportional valve 8 and the elec tromotor 12 or the electromagnetic clutch via signal lines 76.1 , 76.2 or 76.3 are closed to control electronics 78 , which via further signal lines 76 i with further control - Or signal members (not shown) is connected.

The following is the operation of the first embodiment example of the hydraulic clutch betä invention described.

Is a gear change by the driver of the motor vehicle via a jog, preselection or gear lever switch (not shown) manually or by changing one or more control variables such as engine speed, transmission input speed, driving speed, accelerator pedal position etc., via a gear sensor (not automatically initiated, the control electronics 78 starts an automatic coupling process by applying a signal via the signal line 76.3 to the electric motor 12 or the electromagnetic clutch, so that the hydraulic pump 10 is driven by the electric motor 12 or via the electromagnetic clutch is coupled to a rotating shaft of the motor vehicle, for example on a wheel or on the internal combustion engine. From the hydraulic pump 10 , the pressure medium is now largely depressurized starting from the reservoir 14 via the intake line 70 , the hydraulic pump 10 , the pressure line 68 , through the pressure chamber 44 of the proportional valve 8 , past valve body 56 and valve seat 58 through the valve gap 60 in the Drain chamber 42 of the proportional valve 8 and from the drain chamber 42 via the drain line 72 back into the reservoir 14 circulated. Wherein substantially unpressurized To circula- tion of the pressure medium, the piston 22 of the clutch master cylinder 2 against the force of the compression spring 34 of the clutch master cylinder 2 is not beitsraum in the working space 26 in the Ar 26 in moved.

To generate the hydraulic servo pressure and the volume flow in the amplifier chamber 30 of the clutch master cylinder 2 , which are required to lift the clutch pressure plate operatively connected to the clutch slave cylinder 6 , the solenoid 52 of the proportional valve 8 is now controlled by the control electronics 78 via the signal line 76.2 . By means of the magnetic actuator generated by means of armature 48 and magnetic coil 52 , the valve actuator 56 of the proportional valve 8 is actuated by the valve pin 50 of the control piston 46 against the restoring force of the return spring 62 and the inflow force of the circulating pressure medium on the valve body 56 in Pushed towards the valve seat 58 ver. When the valve body 56 approaches the valve seat 58 , the valve gap 60 and thus the flow cross-section for the circulated pressure medium decrease. As a result, a de-defined dynamic pressure is built up in the pressure chamber 44 of the proportional valve 8 in the flow direction of the pressure medium in front of the valve gap 60 , which serves as the servo pressure of the servo circuit S and its size from the width or opening of the valve gap 60 and from the hydraulic pump 10 volume flow depends. At the same time, a reaction force acts on the valve body 56 of the proportional valve 8 , the size of which depends on the effective hydraulic effective area of the valve body 56 when it is in contact with the valve seat 58 . This reac tion force acts via the valve body 56 on the valve pin 50 of the control piston 46 and is in good approximation proportional to the dynamic pressure in the pressure chamber 44 of the proportional valve 8 , ie to the servo pressure of the servo circuit S. Therefore, it is suitable for electromagnetic actuation of the control piston 46 in particular a proportional magnetic drive with a corresponding current-force characteristic, so that the back pressure generated by the pressure medium through the valve gap 60 accumulation pressure or servo pressure of the servo circuit S by stepless adjustment of the valve gap 60 via the valve pin 50 steps and defined adjustable . The viscosity of the pressure means has only a minor influence on the servo pressure in the servo circuit S, since small volume flows of the incompressible pressure medium and its constant circulation via the reservoir 14 ensure sufficient cooling and because of the constant operating temperature in the servo circuit S.

The defined servo pressure of the servo circuit S generated as described above is based on the pressure chamber 44 of the pro portional valve 8 via the pressure line 66 in the booster chamber 30 of the clutch master cylinder 2 and acts on the hydraulic chamber of the piston 22 of the clutch master cylinder 2 , so that the piston 22 is pushed against the force of the compression spring 34 into the working space 26 of the clutch master cylinder 2 . Here, the left-most in Fig. 1 the end of the piston 1 travels over 22-mounted, the working space 26 associated sealing element 24.1, the pressure equalizing and tracking device 18 in Fig. To the left, so that in the Ar beitsraum 26 of the clutch master cylinder 2, a working pressure is built up. This working pressure is on the Drucklei device 4 in the clutch slave cylinder 6 , so that it can lift the clutch pressure plate from the clutch disc.

The displacement sensor 36 arranged on the clutch master cylinder 2 is primarily required for clutch characteristics with a so-called "drop-off" characteristic, in which the disengagement force of the clutch over the disengagement path essentially runs like a parabola open at the bottom, as will be described in more detail below. With the help of the position of the piston 22 of the clutch master cylinder 2 detected by the displacement sensor 36, a clear disengagement path or position assignment of the disengaged clutch pressure plate in relation to the actuating force can then be obtained for the control electronics 78 for control purposes.

If the clutch pressure plate completely disengaged via the clutch slave cylinder 6, the switching operation is at the Ge gear of the motor vehicle by means of the shift lever of the motor vehicle or of the speed encoder.

After completion of the switching process, the clutch pressure plate is brought back into frictional engagement with the clutch disc. For this purpose, according to a predetermined control characteristic, the current applied via the signal line 16 to the magnetic coil 52 is reduced by means of the control electronics 78 , the electromagnetic actuating force on the control piston 46 of the proportional valve 8 being reduced to zero according to the control specification. The dynamic pressure generated in the pressure chamber 44 of the proportional valve 8 by the circulating pressure medium or servo pressure of the servo circuit S acts on the hydraulic effective surface of the valve body 56 and pushes, supported by the restoring force of the return spring 62 , the valve body 56 together with the control piston 46 from valve seat 58 away so that the valve gap 60 increases in the pressure chamber 44 with a decrease of the dynamic pressure, moves until the control piston 46 impact on the on the armature chamber 64 in the fortieth The servo pressure of the pressure medium thus drops to zero at a constant volume flow in the servo circuit S according to a predetermined pressure curve, the piston 22 of the clutch master cylinder 2 , which is acted upon by the hydraulic servo area with its decreasing servo pressure of the servo circuit S, on the basis of the force the compression spring 34 and the hydraulic actuation of its working space-side hydraulic effective surface over the clutch slave cylinder 6 loaded by the resilient clutch pressure plate, is moved into its basic position as far as the stop on the housing base 32 of the clutch master cylinder 2 . The hydraulic connection of the working space 26 of the clutch master cylinder 2 to the reservoir 14 is now via the pressure compensation and trailing device 18 , the trailing line 74 , the drain chamber 42 of the proportional valve 8 and the drain line 72 restored.

The first embodiment of the hydraulic clutch actuation according to the invention was described above in connection with a clutch, the characteristic of which has a "drop-off" characteristic. In Fig. 2A such a characteristic line is shown, as it is in principle with mechanical couplings that are manually disengaged, as a standard application. It is characteristic of this characteristic that the course of the disengagement force F over the disengagement path a is given by constructing the diaphragm springs of the clutch in such a way that both the engagement and disengagement points of the clutch are at a higher force level than the force, which is needed to keep the clutch pressure plate disengaged. This has the advantage that when the clutch pedal is released by depressing the clutch pedal, the actuation force requirement on the clutch pedal is of a measured size, so that even holding the clutch pressure plate in the disengaged state does not tire the driver of the motor vehicle. Because of this characteristic, the displacement sensor 36 is required in the automatic hydraulic clutch actuation described with reference to FIG. 1, in order to enable a clear assignment of the disengagement force F to the disengagement path a and thus to the clutch torque M.

In the automatic hydraulic clutch actuation according to the invention, clutches can also be used which have a clutch characteristic curve with a substantially linear or steadily increasing characteristic, ie with a substantially linear or steadily increasing course of the disengagement force F over the disengagement path a of the clutch pressure plate sen, since the disengagement force F can not be applied by the driver of the motor vehicle and thus the course of the disengagement force F over the disengagement path a is of secondary importance. Such a characteristic is shown in FIG. 2B. The displacement sensor 36 for detecting the disengagement path a of the clutch pressure plate can advantageously be omitted in such a clutch, since the back force F is in a clear connection with the disengagement path a of the clutch pressure plate and thus the clutch torque M.

For a defined or sensitive control of the clutch torque M when engaging the clutch, ie when the clutch pressure plate and the drive plate of the clutch engage via the friction linings, it is sufficient to know the back force F of the clutch or a size associated with it, such as the working pressure in the pressure circuit D, the servo pressure in the servo circuit S, the coil current I _{V of} the magnetic drive in the proportional valve 8 or the current I _{M of} the electric motor 12 to drive the hydraulic pump 10 . The disengagement force F with constant hydraulic effective surfaces on the piston 22 of the clutch master cylinder 2 is only a function of the working pressure in the pressure circuit D or the servo pressure in the servo circuit S, while the working pressure in the pressure circuit D or the servo pressure in the servo circuit S in turn Function of the coil current I _{V of} the magnetic drive in the proportional valve 8 and the current I _{M of} the electric motor 12 for driving the hydraulic pump 10 , there.

Basically, it is possible to detect the working pressure in the pressure circuit D or the servo pressure in the servo circuit S by means of a pressure sensor (not shown) which is connected hydraulically to the pressure line 4 in the pressure circuit D or the pressure line 66 in the servo circuit S and by means of a signal line (not shown) is connected to the control electronics 78 . The pressure sensor supplies actual values for a subordinate pressure control circuit in order to increase the control quality of the overall system and thus to ensure a sensitive engagement of the clutch. With such a scheme, the disengagement of the clutch, ie the separation of the frictional engagement between. Clutch pressure plate and drive plate detected by sensing the speed difference or the slip between the clutch pressure plate and drive plate, and the applied working or servo pressure measured and stored. These values are now available to the control electronics 78 together with further signals which are fed to the control electronics 78 via the signal lines 76 for controlling a fast and comfortable engagement process.

The coil current I _{V of} the magnetic drive in the proportional valve 8 or the current I _{M of} the electric motor 12 for driving the hydraulic pump 10 , which is proportional to the servo pressure in the servo circuit S and is based on the control electronics 78, can also serve as the actual value for such a subordinate control loop the signal line 76.2 or 76.3 is applied to the solenoid 52 in the proportional valve 8 or the electric motor 12 . In this case, a pressure sensor for recording the working or servo pressure is not required. To disengage the clutch when the hydraulic pump 10 is driven by the electric motor 12 , the valve gap 60 of the proportional valve 8 is reduced as described by increasing current supply to the magnetic drive of the proportional valve 8 , so that a servo pressure builds up in the servo circuit S which is applied to the clutch master cylinder 2 and thus an increase in the working pressure in pressure circuit D results. As soon as the disengagement of the clutch, as described above, is detected by speed sensing of the engine and output or clutch pressure plate and drive plate, the coil current I _{V of} the solenoid 52 in the proportional valve 8 or the current consumption I _{M of} the electric motor 12 for driving the hydraulic pump 10 , the now the pressure medium against the resistance of the rising servo pressure in the servo circuit conveys, measures and stores. Based on these values and other signals, the control electronics 78 can now regulate a quick and comfortable engagement process, the engagement of the clutch being effected only by lowering the coil current I _{V of} the solenoid 52 in the proportional valve 8 as described above. As a result, it is possible, when using a clutch with a substantially linear or in any case steadily increasing characteristic of the clutch characteristic line, to regulate the clutch torque M to be transmitted between the clutch pressure plate and the drive plate independently of the return path a of the clutch pressure plate.

Furthermore, in the case of clutches with a characteristic curve according to FIG. 2A, a mixing method from the above-described rules of the engagement process is possible. Accordingly, the engagement of the clutch in the discontinuous portion of the characteristic curve is regulated on the basis of the detected disengagement path or a proportion proportional to it until the course of the characteristic curve over the disengagement path is clear after the apex of the characteristic line has been exceeded. Then, ie in the continuous portion of the characteristic curve, the control is switched over and continued on the basis of the detected disengagement force or a quantity proportional to it, in order to ensure a particularly sensitive engagement of the clutch.

It was described above that the pressure medium before the egg coupling process in the servo circuit S essentially circulates without pressure. An advantageous alternative to the essential The control is provided by the pressureless circulation of the pressure medium tion of the pressure circuit D via the servo circuit S to a moment tracking, the basic idea of which is the dome not up to the maximum transferable clutch torque close, but only close to the extent that the transfer clutch torque always just above the current engine torque ment of the internal combustion engine of the motor vehicle.

In the case of torque tracking, a possible shift request or gear change is recognized by the control electronics 78 at an early stage by changing one or more control parameters, such as engine speed, driving speed, accelerator pedal position, which are supplied to the control electronics 78 via the signal lines 76 i then controls the pressure circuit D via the servo circuit S as described above to generate a clutch torque via the clutch slave cylinder 6 on the clutch, that is always just above the current engine torque.

One advantage of moment tracking is especially in that see that the clutch pressure plate with reduced Stellge  dizziness in time even when changing gears quickly can be completely disengaged.

Also, the torque tracking unfolds an advantageous effect on load changes, the vibrations easily generated on the clutch or induced in the clutch actuation by quickly changing the engine torque when the clutch is closed, so that due to the slip occurring only briefly on the clutch Consumption and wear advantages result. Here, the clutch torque is adjusted with a predetermined hysteresis, that is, the clutch torque is only changed when the engine torque has changed by a certain minimum amount, which results in a reduction in the on-time of the electric motor 12 or the electromagnetic clutch Drive of the hydraulic pump 10 is connected.

Fig. 3 shows a variant of the first embodiment shown in Fig. 1 of the hydraulic clutch actuation according to the invention. The variant according to FIG. 3 differs from the first exemplary embodiment shown in FIG. 1 essentially in that another proportional valve 8 'in the servo circuit S is used. In the fol lowing the variant of FIG. 3 is described only in terms of its differing features from the first embodiment of FIG. 1.

In the proportional valve 8 'shown in Fig. 3', which is a switching variant of the proportional valve 8 according to FIG. 1, it is an electromagnetically actuated 2/2 ball seat valve which is switched in the non-actuated state in the blocking zero position. The proportional valve 8 'has a three-chamber valve housing 38 ', which defines an armature chamber 40 ', an outlet chamber 42 ' and a pressure chamber 44 '. The armature chamber 40 ', the pressure chamber 44 ' and the drain chamber 42 'are arranged in this order in Fig. 3 from right to left one behind the other concentrically to the central axis of the valve housing 38 '. The drain chamber 42 'is connected to the reservoir 14 via the drain line 72 and to the pressure compensation and follow-up device 18 in the clutch master cylinder 2 via the follow-up line 74 , while the pressure chamber 44 ' via the pressure line 68 to the hydraulic pump 10 and via the pressure line 66 the booster chamber 30 in the clutch master cylinder 2 is connected.

In the armature chamber 40 'of the proportional valve 8 ', a control piston 46 'is received, which consists of a ferromagnetic armature 48 ' with a valve pin 50 'arranged centrally thereon and in the armature chamber 40 ' is axially displaceable electromagnetically. For this purpose, a magnetic coil 52 'is provided in the radial wall of the armature chamber 40 ', which surrounds the armature 48 'at least partially concentrically and together with the armature 48 ' forms a magnetic drive with the aid of which the control piston 46 'in the axial direction of the armature chamber 40 ′ Can optionally be moved. The valve bolt 50 'of the control piston 46 ' is in a partition 54 'of the valve housing 38 ' between the armature chamber 40 'and the pressure chamber 44 ' hy cally tight and protrudes into the pressure chamber 44 '.

The pressure chamber 44 'is connected to the drain chamber 42 ' through an axis with the control piston 46 'lying through or stepped bore, which is provided on the discharge chamber side with a centrally worked valve seat 58 '.

In the drain chamber 42 'is a valve body 56 ' is arranged, which is designed as a metallic ball, which is sealed by the partition wall 54 'between the armature chamber 40 ' and the pressure chamber 44 'extending valve bolt 50 ' of the control piston 46 'mechanically with a Actuating force can be applied. The valve body 56 'can only be acted upon by a pressure force via the valve pin 50 '. The stroke of the armature 48 'of the control piston 46 ' in the armature chamber 40 'corresponds to the required valve opening path of the valve body 56 ' in the drain chamber 42 '.

In the drain chamber 42 'is also a closing spring 63 is arranged, which presses the valve body 56 ' in its basic position against the valve seat 58 ', whereby in the locked zero position of the proportional valve 8 ' the pressure chamber 44 'and the drain chamber 42 ' are hydraulically tightly separated. The closing spring 63 is dimensioned such that the amount of its biasing force is equal to the amount of force acting on the valve body 56 'pressure chamber side when the maximum working pressure required to disengage the clutch in the pressure circuit D is present. For a clutch characteristic with "drop-off" characteristic, the maximum working pressure required to disengage the clutch is the pressure required to generate the force at the apex of the disengagement force curve F in FIG. 2A, while the maximum working pressure required to disengage the clutch in the case of a clutch characteristic curve with a continuously increasing characteristic, that pressure which is required to generate the release force F at which no more moment M is transmitted (dashed line in FIG. 2B).

In operation, the hydraulic clutch actuation in accordance with Fig. 3 needs the magnetic coil 52 'of the proportional valve 8' are thus not energized to disengage the clutch, since the valve body 56 'against the force of the closing spring 63 until the valve seat 58' can take off, if the Hydraulic pump 10, a servo pressure is generated in the servo circuit S, which is above the servo pressure required to generate the maximum required working pressure in the pressure circuit D. To engage the clutch, the solenoid 52 'is energized so that via the valve pin 50 ' of the control piston 46 'a defined Ven valve gap between the valve body 56 ' and the valve seat 58 'can be set, by means of which the servo pressure in the ser vokreis S as below Described with reference to FIG. 1 is controllable. The regulation of the engagement process can depend in principle on the clutch characteristic as described with reference to FIG. 1, with the restriction that with a clutch characteristic with steadily increasing characteristics of the coil current I _{V of} the magnetic drive in the proportional valve 8 'not as Actual value can serve for the control circuit, since the magnetic drive is not energized when the clutch is disengaged.

FIG. 4 shows the second embodiment of the OF INVENTION hydraulic clutch actuation to the invention in the unactuated state. The second embodiment differs from the first embodiment according to FIG. 1 only in that no clutch master cylinder is provided and thus a separate pressure circuit is omitted. According to this embodiment, the servo circuit S without trailing line 74 via the pressure line 66 is directly connected to the clutch slave cylinder 6 , which is provided with the displacement sensor 36 . Structure and function of the servo circuit S otherwise correspond to that of the first exemplary embodiment from FIG. 1, so that at this point a further description can be omitted. Also be described with reference to the first embodiment, the method for regulating the disengagement and engagement of the clutch with the second embodiment can in principle be carried out accordingly. Finally, the servo circuit S of the second embodiment can also be constructed as shown in FIG. 3, the pressure circuit D with clutch master cylinder 2 and the follow-up line 74 being omitted and the pressure line 66 being connected directly to the clutch slave cylinder 6 .

FIG. 5 shows the third embodiment of to the invention OF INVENTION hydraulic clutch actuation in the unactuated state. Parts corresponding to the parts in FIG. 1 are provided with the same reference symbols and are not explained again below. Although the third embodiment in terms of structure and function of the servo circuit S and the pressure circuit D essentially corresponds to the first embodiment according to FIG. 1, depending on the application, the servo circuit S can also be constructed as shown in FIG. 3 or only the servo circuit S according to Fig. 4 may be provided. In all cases, however, in the pressure line 68 between the hydraulic likpump 10 and the proportional valve 8 , 8 'to the hydraulic likpump 10 blocking check valve is arranged, which is not shown in Fig. 5 for reasons of clarity.

For an automatic switching process, which is preceded by the automatic clutch release process, the circuit S is connected according to FIG. 5 to a hydraulic circuit x which has a plurality of switching or gear stages A 1 to A _Z. The structure and function of only the first switching stage A 1 in the circuit X is described below. Training, order and function of the other switching stages A₂ to A _Z in the circuit X, the number Z depends on the transmission design, are identical for each switching stage and therefore do not need to be explained in more detail.

Referring to FIG. 5, the first shift stage A₁ is provided with a further proportional valve 80, which is an elec-magnetically actuated 2/2 ball seat valve which is connected in the zero position locking in the unactuated state in. This proportional valve 80 represents a shift variant of the proportional valve 8 of the servo circuit S and controls a shift cylinder 82 for actuating a sliding sleeve 84 (not shown in FIG. 5) on a hub (not shown) in the transmission. A preferred embodiment of the sliding sleeve 84 is described in more detail with reference to FIG. 7.

The shift cylinder 82 has a piston 86 , which divides a cylinder bore 88 of the shift cylinder 82 into a working space 90 and a rear compartment 92 . In the basic division of the shift cylinder 82 , the piston 86 strikes the cylinder bottom 94 of the shift cylinder 82 . As already mentioned, each gear or shift stage A 1 to A _{Z is provided} with a proportional valve 80 and a shift cylinder 82 .

The circuit X is also ver with a switching valve 96 in the form of an electromagnetically controllable 2/2-way valve, which is switched into the blocking position as the basic position and through position when actuated and whose circuitry with the other components of the circuit X is described in more detail. Only one switching valve 96 is required in circuit X.

The proportional valve 80 shown enlarged in FIG. 6 has a three-chamber valve housing 98 , which delimits an armature chamber 100 , a control chamber 102 and a pressure chamber 104 . The armature chamber 100 , the control chamber 102 and the pressure chamber 104 are arranged in this order in FIG. 6 one behind the other concentrically to the central axis of the valve housing 98 .

In the armature chamber 100 of the proportional valve 80 , a control piston 106 is accommodated, which consists of a ferromagnetic armature 108 with a valve bolt 110 arranged centrally thereon and is axially displaceable electromagnetically in the armature chamber 100 . For this purpose, a magnetic coil 112 is provided in the radial wall of the armature chamber 100 , which surrounds the armature 108 at least partially concentrically and forms a magnetic drive together with the armature 108 , with the aid of which the control piston 106 can optionally be moved in the axial direction of the armature chamber 100 . The valve pin 110 of the control piston 106 is in a partition 114 of the valve housing 98 between the armature chamber 100 and the control chamber 102 hy cally tight and protrudes into the control chamber 102 .

The control chamber 102 is connected to the pressure chamber 104 by a through or stepped bore 116 lying on one axis with the control piston 106, which is provided on the pressure chamber side with a centrally worked valve seat 118 .

In the pressure chamber 104, a valve body 120 is arranged, which is formed as a metallic ball which is sealed by the partition wall 114 between the armature chamber 100 and the control chamber 102 therethrough Ventilbol zen 110 of the control piston 106 mechanically purification capacity with a Actuate the can be applied over the. The valve body 120 can only be subjected to a compressive force via the valve pin 110 of the control piston 106 , since the valve body 120 and the valve pin 110 are two separate components. The stroke of the armature 108 of the control piston 106 in the armature chamber 100 corresponds to the required valve opening path of the valve body 120 which is arranged in the pressure chamber 104 and is operatively connected to the valve pin 110 of the control piston 106 .

In the pressure chamber 104 , a return spring 122 is also arranged, which presses the valve body 120 in its basic position against the valve seat 118 , whereby the control chamber 102 and the pressure chamber 104 are hydraulically tightly separated from one another in the zero position of the proportional valve 80 . The return spring 122 also pushes the valve piston 106 lying in series with the valve body 120 to a stop 124 which is formed on the valve housing 98 at the end of the armature chamber 100 facing away from the control chamber 102 .

In the following, the United circuit of the proportional valve 80 with the switching cylinder 82 and the switching valve 96 in the circuit X and the circuitry of the circuit X with the servo circuit S will be described with reference to FIG. 5.

The pressure chamber 104 of the proportional valve 80 is connected via a pressure line 126 to the pressure line 66 between the Druckkam mer 44 of the proportional valve 8 in the servo circuit S and the United amplifier chamber 30 of the clutch master cylinder 2 . For the basic position of the first switching stage A 1 in the circuit X, the pressure line 126 branches into a pressure line 128 which connects the reset space 92 of the switching cylinder 82 and the pressure chamber 104 of the proportional valve 8 to one another. Wei terhin is connected between the proportional valve 80 and the switching cylinder 82, a control line 130 which connects the control chamber 102 of the proportional valve 80 with the working space 90 of the switching cylinder 82 . The control line 130 is connected via a control line 132 to a manifold 134 , which is connected via the switching valve 96 to the suction line 70 between the reservoir 14 and the hydraulic pump 10 , so that the hydraulic circuit is closed. As a further res functional element, a check valve 136 with a blocking effect on the control line 130 or passage to the collecting line 134 is provided in the control line 132 .

Finally, the solenoid 112 of the proportional valve 80 is connected via a signal line 76.4 to the control electronics 78 , while the switching valve 96 is connected by means of a signal line 76.5 to the control electronics 78 .

The following is the operation of the third embodiment example of the hydraulic clutch actuation according to the invention described.

If after the coupling process, which was carried out as described with reference to FIGS. 1, 3 and 4, the automatic switching process is carried out, then in the basic position shown in FIG. 1 the first switching stage A 1 with servo-pressurized clutch master cylinder 2 and by the Clutch slave cylinder 6 lifted clutch pressure plate over the pressure lines 66 , 126 and 128 in the back space 92 of the shift cylinder 82 and over the pressure lines 66 and 126 in the pressure chamber 104 of the proportional valve 80 of the servo pressure, which holds the piston 86 at the stop on the cylinder bottom 94 and the closing force of the return spring 122 spring-loaded valve body 120 of the proportional valve 80 to the valve seat 118 is additionally increased. The Druckmit tel in the control chamber 102 of the proportional valve 80 , the control lines 130 and 132 , the manifold 134 and the Ar beitsraum 90 of the switching cylinder 82 is pressure-free ruled out. The proportional valve 80 and the switching valve 96 are in their blocking basic position.

Via the control electronics 78 , the solenoid 112 of the proportional valve 80 of the required switching stage A₁, A₂, A _Z is now driven, whereupon the control piston 106 of the proportional valve 80 is electromagnetically driven in the direction of the valve body 120 and moves against the mechanical force of the return spring 122 as well as the hydraulic closing force on the valve body 120 pushes from the valve seat 118 , so that a valve gap opens between the valve seat 118 and the valve body 120 . As a result, a switching pressure is generated when flowing through the valve gap in the proportional valve 80 , which propagates via the control chamber 102 of the proportional valve 80 into the control lines 130 and 132 , the manifold 134 and the working space 90 of the Schaltzyliners 82 .

The piston 86 of the shift cylinder 82 , which is operatively connected to the sliding sleeve 84 to be described with reference to FIG. 7, is acted upon by a force which is the product of the difference between the hydraulic active surfaces on both sides of the piston 86 and corresponds to the pending switching pressure in the reset space 92 , and moved ben, so that the shifting operation is carried out under synchronization with the sliding sleeve 84 .

At the end of the switching process, the electrical control of the solenoid 52 of the proportional valve 8 and the solenoid 112 of the proportional valve 80 is ended via the control electronics 78 and the switching valve 96 is switched to the zero position. As a result, no servo pressure and thus no switching pressure is generated. The piston 22 of the clutch master cylinder 2 moves as described with reference to FIG. 1 in its basic position, so that the clutch pressure plate comes back into engagement with the clutch disc. The valve body 120 is pushed by the force of the return spring 122 on the valve seat 118 and the switching valve 96 is switched to the locked zero position. The piston 86 of the shift cylinder 82 is thus held in its forward or extended position by the pressure medium enclosed between the control chamber 102 of the proportional valve 80 and the Schaltven valve 96 . The switching stage is now engaged.

For automatic downshifting of the inserted gear stage into idle while the internal combustion engine is running or for changing to another gear stage, the switching valve 96 is required, the automatic coupling process first being carried out as described with reference to FIGS . 1, 3 and 4 and then that Electromagnetically operated switching valve 96 is switched from its blocking zero position to its through position. With the aid of pressure on the line 128 applied pressure in the working chamber 90 of the Schaltzy Linders 82 of the piston 86 of the shift cylinder 82 together with the sliding sleeve 84 derboden 94 by the pressure medium to stop at Zylin pushed back of the shift cylinder 82nd The previously locked in the working space 90 of the switching cylinder 82 pressure medium is pushed through the control lines 130 , 132 through the check valve 136 and via the manifold 134 through the GE opened switching valve 96 into the intake line 70 . With the stop of the piston 86 on the cylinder bottom 94 of the Schaltzylin ders 82 , the switching valve 96 is switched back to its basic position. The switching stage A₁ is now uncoupled.

For a subsequent change to another gear level the coupling process and the switching process as explained above to perform accordingly.

In Fig. 7, a preferred embodiment of the connection of the double-acting shift cylinder 82 , 82 'two switching stages with the above-mentioned sliding sleeve 84 of the transmission is provided. According to FIG. 7, the piston 86 of the shift cylinder 82 is firmly connected to a shifting claw 138 which is only shown schematically and which engages in a form-fitting manner in a recess 140 provided on the outer circumference of the sliding sleeve 84 . The grip contour of the switching claw 138 can be designed, for example, as a simple hook or as a fork with radially offset engagement points.

The sliding sleeve 84 is arranged axially displaceable on both sides by a path that corresponds to the stroke h of the shift cylinder 82 from a central position along the central axis of the shift cylinder 82 . The path of the sliding sleeve 84 for a switching stage is composed of an idle travel l 1 and a shift travel l 2. On the opposite side of the sliding sleeve 84 and depending on the structural design angularly offset to the shift cylinder 82 is the shift cylinder 82 'for the next shift stage, at play, the switching stage A₂ arranged on the circumference of the sliding sleeve 84th

In its basic position, the sliding sleeve 84 is hydraulically in the manner described above by the pistons 86 , 86 'positioned on the cylinder bottoms 94 , 94 ' on the stop, the shift cylinders 82 , 82 'and the positively engaging in the recess 140 of the sliding sleeve 84 shifting claws 138th , 138 ′ arrested.

If now during a switching operation the pressurized piston 86 of the shift cylinder 82 and thus the sliding sleeve 84 is axially displaced on the hub of the transmission, the shifting claw 138 arrives at the opposite side of the wall of the recess 140 of the sliding sleeve 84 and moves after crossing the free travel l 1 the sliding sleeve 84 around the switching path l₂. The shifted position of the sliding sleeve 84 is shown in FIG. 7 in dashed lines. The shift claw 138 'of the shift cylinder 82 ', which remains in its basic position, of the next shift stage remains disengaged from the sliding sleeve 84 . This is due to the fact that the thickness s of the respective switching claw 138 , 138 'is smaller than the free travel l 1 of the sliding sleeve 84 .

When changing to the next switching stage, as described above, first the piston 86 of the switching cylinder 82 moves together with the sliding sleeve 84 into the basic position, according to which the procedure described with reference to FIGS . 1, 3 or 4 and 5, for example, the Switching stage A₂ with the switching cylinder 82 'can be switched.

A method and a device for hydraulically actuating a clutch operatively connected to a clutch slave cylinder, in particular for motor vehicles, is disclosed, in which a servo pressure is generated in a servo circuit, which
the hydraulic control of a clutch master cylinder serves to generate a working pressure in a pressure circuit, which is applied to the clutch slave cylinder, or
is applied directly to the clutch slave cylinder in order to apply the force required to disengage the clutch. According to the servo pressure for disengaging and disengaging the clutch in the servo circuit is set in a defined manner, for which purpose the servo circuit has a proportional valve. As a result, automated actuation with improved response and operating behavior of the clutch is made possible with little expenditure on device technology. Furthermore, a method and an apparatus for switching a mechanical transmission bes is disclosed, in which the servo pressure of the clutch disengagement is used to switch the sliding sleeves of the mechanical transmission.

Reference list

2 clutch master cylinders
4 pressure line
6 clutch slave cylinders
8 proportional valve ( 8 ′)
10 hydraulic pump
12 electric motor
14 storage containers
16 housing
18 Pressure equalization and follow-up device
20 housing bore
22 pistons
24.1 sealing element
24.2 Sealing element
26 work space
28 trailing space
30 amplifier room
32 case back
34 compression spring
36 displacement sensors
38 valve housing ( 38 ′)
40 anchor chamber ( 40 ′)
42 drain chamber ( 42 ′)
44 pressure chamber ( 44 ′)
46 control pistons ( 46 ′)
48 anchors ( 48 ′)
50 valve bolts ( 50 ′)
52 solenoid ( 52 ′)
54 partition ( 54 ′)
56 valve body ( 56 ′)
58 valve seat ( 58 ′)
60 valve gap
62 return spring
63 closing spring
64 stop
66 pressure line
68 pressure line
70 suction line
72 drain pipe
74 Follow-up pipe
76 signal line ( 76.1-76.5 , 76 i)
78 Control electronics
80 proportional valve
82 shift cylinders ( 82 ′)
84 sliding sleeve
86 pistons
88 cylinder bore
90 work space
92 storage room
94 cylinder base ( 94 ′)
96 switching valve
98 valve housing
100 anchor chamber
102 control chamber
104 pressure chamber
106 spool
108 anchors
110 valve bolts
112 solenoid
114 partition
116 stepped bore
118 valve seat
120 valve body
122 return spring
124 stop
126 pressure line
128 pressure line
130 control line
132 control line
134 manifold
136 check valve
138 shifting claw ( 138 ′)
140 puncture
A₁ first switching stage
A₂ second switching stage
A _Z last switching stage
D pressure circuit
F release force
I _M Current of the electric motor for driving the hydraulic pump
I _V coil current of the solenoid actuator in the proportional valve
M clutch torque
S servo circuit
X circuit
a Release path of the clutch pressure plate
h shift cylinder stroke
l₁ free travel of the sliding sleeve
l₂ switching path of the sliding sleeve
s thickness of the shifting claw

Claims (18)

1. A method for hydraulic actuation of a clutch slave cylinder ( 6 ) operatively connected clutch, in particular for motor vehicles, according to which a servo pressure is generated in a servo circuit (S)
the hydraulic control of a clutch master cylinder ( 2 ) is used to generate a working pressure in a pressure circuit (D) which is applied to the clutch slave cylinder ( 6 ), or
is applied directly to the clutch slave cylinder ( 6 ) in order to apply a force required to disengage the clutch to the clutch slave cylinder ( 6 ), characterized in that the servo pressure for disengaging and engaging the clutch in the servo circuit (S) is set in a defined manner. 2. The method according to claim 1, characterized in that the Disengaging a clutch, the characteristic of the disengagement force (F) over the disengagement path (a) essentially like one down open parabola runs, the release path (a) or one to it proportional size along with a slip between Input and output of the clutch representing the size summarizes and engages the clutch based on the detected Sizes is regulated. 3. The method according to claim 2, characterized in that the position of the piston ( 22 ) in the clutch master cylinder ( 2 ) or the position of the piston in the clutch slave cylinder ( 6 ) serves as the proportional to the release path (a) of the clutch. 4. The method according to claim 1, characterized in that the Disengaging a clutch, the characteristic of the disengagement force (F) steadily over the release path (a) with increasing release path (a) increases, the disengagement force (F) or a proportional Size together with a the slip between drive and down drive the size representing the clutch detected and the on  back clutch regulated on the basis of the detected sizes becomes. 5. The method according to claim 4, characterized in that the Working pressure in the pressure circuit (D) or the servo pressure in the servo circle (S) as the proportio to the disengagement force (F) of the clutch size serves. 6. The method according to claim 4, characterized in that the coil current (I _V ) of a magnetic drive of a proportional valve ( 8 ) provided in the servo circuit (S) for regulating the servo pressure or the coil current (I _M ) of a drive for a hydraulic pump ( 10 ) the electric motor ( 12 ) provided in the servo circuit (S) serves as the variable proportional to the release force (F) of the clutch. 7. The method according to claim 1, characterized in that the Engaging a clutch, the characteristic of the disengagement force (F) over the disengagement path (a) essentially like one down The open parabola runs in the discontinuous section of the Characteristic curve according to the method according to claim 2 or 3 and in the continuous section of the characteristic curve according to the method is regulated according to one of the claims 4 to 6. 8. The method according to any one of the preceding claims, characterized in that the engine torque of a drive connected to the clutch is determined and a servo pressure is set as a function of the average engine torque in the servo circuit (S), which cylinder directly or via the clutch master ( 2 ) of the clutch slave cylinder ( 6 ) applies a force to the clutch which enables the transmission of such a torque through the clutch, which is smaller than the maximum transmissible torque by the clutch, but is larger by a predetermined amount than the currently determined engine torque. 9. The method according to claim 8, characterized in that the First determine in the servo circuit (S) depending on one Only then do the servo pressure set again is set when a later determined engine torque a predetermined amount from the initially determined engine torque deviates. 10. Device for performing the method according to one of the preceding claims, with a servo circuit (S), in which a servo pressure can be generated, which is applied to a servo space ( 30 ) of a clutch master cylinder ( 2 ) in order to act upon a piston ( 22 ) to generate a working pressure in a working space ( 26 ) of the clutch master cylinder ( 2 ) which is applied to a clutch slave cylinder ( 6 ) which is operatively connected to the clutch in order to apply a disengagement force, characterized in that the servo circuit (S) has a proportional valve ( 8 , 8 '), by means of which the servo pressure in the servo circuit (S) is adjustable. 11. The device according to claim 10, characterized in that the servo space hydraulic effective area of the piston ( 22 ) of the clutch master cylinder ( 2 ) is larger than the working area hydraulic effective area of the piston ( 22 ). 12. An apparatus for performing the method according to one of claims 1 to 9, with a servo circuit (S) in which a servo pressure can be generated, which is applied to apply a disengagement force on the clutch to a clutch slave cylinder ( 6 ) which is operatively connected to the clutch , characterized in that the servo circuit (S) has a proportional valve ( 8 , 8 '), by means of which the servo pressure in the servo circuit (S) is de finely adjustable. 13. Device according to one of claims 10 to 12, characterized in that the proportional valve ( 8 ) has a valve body ( 56 ) which limits a valve gap ( 60 ) which can be positively flowed through when the clutch is actuated by a pressure medium, so that the servo pressure can be controlled as electromagnetically Back pressure in a pressure chamber ( 44 ) of the proportional valve ( 8 ) to be defined, which is hydraulically connected to the servo space ( 30 ) of the clutch master cylinder ( 2 ) or the clutch slave cylinder ( 6 ). 14. The apparatus according to claim 13, characterized in that the pressure chamber ( 44 ) of the proportional valve ( 8 ) to a hy draulic pump ( 10 ) is connected, the pressure medium through the valve gap ( 60 ) of the proportional valve ( 8 ) when the clutch is actuated. promotes. 15. The device according to any one of claims 10 to 12, characterized in that the servo circuit (S) has a proportional valve ( 8 ') biased in blocking-zero position, which one with a pressure source ( 10 ) and the servo space ( 30 ) of the clutch master cylinder ( 2 ) or the clutch slave cylinder ( 6 ) hy draulically connected pressure chamber ( 44 ') and a valve body ( 56 '), which limits an electromagnetically controllable valve gap when the clutch engages with a pressure medium, to the servo pressure as dynamic pressure in the pressure chamber ( 44 ') defined, and which is biased against the pressure in the pressure chamber ( 44 ') by means of a closing spring ( 63 ), the biasing force is equal to the force acting on the valve body ( 56 ') pressure chamber side when the the maximum required working pressure in the pressure circuit (D) or servo pressure in the servo circuit (S) is present to disengage the clutch. 16. The device according to one of claims 13 to 15, characterized in that the proportional valve ( 8 ; 8 ') is formed as a ball seat valve with a spherical valve body ( 56 ; 56 '), which together with an annular valve seat ( 58 ; 58 ') Limits the valve gap ( 60 ). 17. Device according to one of claims 10 to 16, characterized in that a circuit (X) is connected to the servo circuit (S), which serves for the hydraulic actuation of a manual transmission, for which purpose the circuit (X) for each switching stage (A₁, A₂ , a _Z) of the gearbox a Schaltzylin of (82) and an attached to a pressure side of the servo loop (S) connected proportional valve (80) for driving the switching cylinder (82), and a switching valve (96) that the circuit (X ) optionally connects to a suction side of the servo circuit (S). 18. The apparatus according to claim 17, characterized in that the proportional valve ( 80 ) of the circuit (X) is an electro-magnetically controllable ball seat valve, which is biased in the locked zero position.