Abstract:
A friction clutch for use in a gear changing transmission for facilitating a power transmission from a rotating shaft of a working exhaust-gas turbine onto a crankshaft of an internal combustion engine, which friction clutch, depending on the engine and/or turbine speed, automatically closes in the direction of engagement and opens in the direction of disengagement, with the friction members of the clutch being able to be compressed in the direction of engagement against a spring force (cup springs) by a piston of a rotating cylinder-piston-unit provided on the drive side of the clutch, such that a fluid conveyed from an external pump into the cylinder applies a fluid pressure onto the piston in response to rotation (centrifugal force). The engine lubricant is preferably used as the fluid.

Description:
FIELD OF THE INVENTION 
     The invention relates to a friction clutch having friction members therein for use in a gear changing transmission for facilitating a power transmission from a working exhaust-gas turbine to a crankshaft of an internal combustion engine. Such turbines are used in commercial vehicles, like trucks and the like to conserve fuel. 
     BACKGROUND OF THE INVENTION 
     Pressure-medium operated friction clutches, as they are known in various designs, cannot be used for the above-mentioned purpose, because the starting and turning-off operations must occur independently dependent on the turbine and/or engine speed in order to avoid additional control devices. Also mechanically shiftable friction clutches, which meet the demands of the specific application are not known. 
     Therefore the basic purpose of the invention is to provide a clutch for the purpose mentioned above. 
     The purpose is attained by providing a friction clutch having friction members therein which, depending on the speed of the engine and/or the speed of the turbine, automatically closes in the direction of engagement and opens in the direction of disengagement, with the friction members being compressible in direction of engagement against the force of a spring member by a piston of a rotating cylinder-piston-unit provided on a drive side of the clutch, a fluid being provided and conveyed by an external pumping device into a cylinder applying in response to rotation a pressure onto the piston. That is, the supplied fluid applies a pressure onto the piston under the action of centrifugal force in the cylinder rotating dependent on the turbine speed, which pressure compresses the friction members against the force of the spring. When the turbine speed drops and the pressure of the fluid decreases, the clutch is again opened by the spring. 
     An advantageous embodiment is where the cylinder is connected to a driving element, namely, the output shaft of the turbine and is designed with an annular chamber therein for receiving the piston therein and is supported rotatably on a shaft arranged on a driven side of the clutch. Using the motor lubricant as the fluid for the clutch is very advantageous because no additional closed fluid circuit is needed for the operation of the clutch. The pump, which is provided anyway, supplies also the clutch with oil. Since the pressure of the supplied engine lubricant is not constant and since the cylinder has only a limited volume in which a pressure dependent on the turbine speed is to build up, a relief bore is provided through which all oil not needed for the pressure build-up is discharged or through which a lubricant pressure, which may possibly be too high, is reduced. 
     The piston is advantageously secured against an unintentional rotation. A limited relative rotation between the piston and the cylinder can thereby be utilized to interrupt the fluid supply in response to the speed on the driven side of the clutch being higher than the speed on the driving side of the clutch. This provides an overload safety measure for the turbine should there occur a sudden speed increase of the engine, for example, during down-shifting on an incline. An advantageous further development of the clutch includes the provision of at least one curved groove which is covered toward the annular chamber by a disk or the like resiliently resting on the piston. The resilient cover of the groove permits the fluid in the groove to escape only slowly during the relative rotation between piston and cylinder, which avoids a sudden opening and closing of the clutch. 
     A further development of the clutch which includes a device for interrupting the fluid supply at a low speed of said engine (idling speed) is sensible in order to uncouple the turbine from the engine during idling of the engine, even if the clutch is still closed. This is particularly sensible during a sudden speed reduction of the engine, for example, when the vehicle driver&#39;s foot slips off from the clutch pedal. An advantageous embodiment for a device for effecting an uncoupling of the turbine from the engine in such cases is through an interruption of the fluid supply to the clutch. Since a 100% seal cannot be achieved with simple structure, a further relief bore is advisable. Since only leakage oil needs to be discharged and the pressure build-up in the cylinder must be assured during normal fluid supply, this bore is to have only a very small cross section. 
     To protect the clutch against vibrations (rotational irregularities) coming from the engine, which vibrations are very high particularly in the lower speed range of the engine, an attenuating or damping member is provided between the clutch and a crankshaft of the engine. The attenuating or damping member (vibration attenuator) does not need to be integrated into the clutch, but can also be arranged separately from the clutch, for example, within a transmission step or between two transmission steps, which is or are arranged between the clutch and the crankshaft of the engine. 
     To achieve a satisfactory start of the closing operation without any delays, a further development includes the provision of plural friction members, each of which include friction surfaces which rest with little pressure on one another, also when said clutch is disengaged. One very simple design includes at least one spring for pressing the disk against the piston to cause initial contact of the friction surfaces. To limit the maximum transmittable torque, the path covered by the piston during engagement is limited by a stop. 
    
    
     BRIEF DESCRIPTION OF THE DRAWINGS 
     The invention will be described hereinbelow with reference to one exemplary embodiment illustrated in the drawings, in which: 
     FIG. 1 illustrates a schematic diagram of a transmission into which a clutch embodying the invention is inserted; 
     FIG. 2 is a cross-sectional view of such a transmission with a slightly modified design; 
     FIG. 3 is a partial cross-sectional view taken along line III--III of FIG. 2; and 
     FIG. 4 is a partial cross-sectional view taken along the line IV--IV of FIG. 2. 
    
    
     DETAILED DESCRIPTION 
     A planetary gearing 2, a friction clutch 3 (hereinafter referred to only as a clutch 3) and a vibration attenuator or damper 4 are series connected and housed in a transmission housing 1. A drive occurs from a turbine shaft 5 of a working exhaust-gas turbine 6 onto a sun gear 7 of the planetary gearing 2 designed with stationary planet gears 8. The planetary gearing 2 through planet gears 8 effect a rotating of a ring gear 9 and the clutch 3 connected to it. The clutch 3, which will be discussed in greater detail later on, couples and uncouples the turbine shaft 5 to and from, respectively, the crankshaft 15 of an internal combustion engine 10. Gears 11, 12, 13, 14 are arranged between the driven part of the clutch 3 and the engine 10 for the purpose of facilitating a speed adjustment. The gears 11, 12, 13 and 14 are housed in a separate housing (not illustrated). The vibration attenuator or damper 4 is provided on the output side of the clutch 3 immediately following the clutch. 
     Details of the clutch 3 or the transmission are illustrated in FIG. 2, where corresponding parts have the same reference numerals as in FIG. 1. The ring gear 9 of the planetary gearing 2 is supported on a shaft 20 through a ring-gear carrier 19, to which it is connected fixedly and pressure-tight. The shaft 20 is rotatably supported in the transmission housing 1 and in a planet gear carrier 18. The planet gear carrier 18 is, for structural reasons, secured as by bolts 17 to the housing 16 of the working exhaust-gas turbine. The carrier 18 could, however, also be secured directly to the transmission housing 1. 
     The ring gear 9 with the ring-gear carrier 19 forms a cylinder 21, in which a ring piston 22 is movably guided, however, pressure-tight The sealing elements do not need to be illustrated and described, since they are known. A snap ring 23 in the cylinder limits the path over which the ring piston 22 can be moved. An annular shoulder on the ring piston 22 has notches 24 for receiving a nose part of so-called external disks 25. The latter cooperate with so-called internal disks 26, which engage with an internal tooth system a corresponding profile in a sleeve-shaped shoulder 27 on the gear 11. The so formed disk package abuts a pressure plate 28 on a side remote from the ring piston 22. The pressure plate 28 is urged by cup springs 29 in the direction of clutch engagement. The cup springs 29 are supported on the gear 11 and press the pressure plate 28 against a snap ring 30 arranged such that the disks are not compressed by the pressure plate 28 or by the cup springs 29 when the clutch is open. 
     Fluid is continuously fed from an external pump P to an annular chamber 31 formed by the cylinder 21 and between the ring piston 22 ring gear carrier 19. The fluid is for simplicity reasons a lubricant of the engine 10 which is fed by the lubricant pump P of the engine through not illustrated pipelines to the transmission housing 1. Bores 32, 33 are for this purpose provided in the housing 1, through which bores 32, 33 oil is fed to a central bore 34 in the shaft 20 and thence on through radial bores 35 in the shaft 20, 36 in a sleeve 37 and 38 in a wall 21A of the cylinder 21 and through a slot 39 on a side of the ring piston 22 remote from the clutch 3 into the annular chamber 31. 
     Through a measured lubricant feed and suitable dimensioning of the piston surface a pressure builds up in the annular chamber 31 dependent on the speed of the turbine by rotating the cylinder filled with the lubricant as fluid. The pressure increases at a squared exponential rate with respect to the rotative speed until the retaining force of the cup springs 29 is exceeded and thus the disk package 25, 26 is compressed. The torque emitted by the turbine is in this manner transmitted through the gears 11, 12, 13, 14 onto the crankshaft of the engine 10. The transmittable torque depends on the number of friction surfaces, the friction value of the disks and the contact pressure of the piston, which as a variable depends on the speed of the turbine. 
     The transmittable torque resulting from these three values must always be greater than the torque output from the turbine. However, torque peaks coming from the side of the engine may not result in an overloading of the turbine. This is particularly valid for maximum turbine speed, where the torque to be transmitted by the clutch reaches its maximum value. In order to limit also in this point of operation the relationship between the torque transmittable from the clutch and the torque emitted by the turbine to a value nondamaging for the entire aggregate, in particular, however, for the turbine, the contact pressure of the piston 22 is limited by the initially pretensioned cup springs 29, which upon reaching the initial tension give way suitably to the piston 22 until the piston rests on the snap ring 23. The now existing initial tension can no longer be exceeded, since a further increase of the pressure in the annular chamber 31 would result in a relative slipping of the disks 25, 26. 
     To build up the pressure in the annular chamber 31, an amount of oil corresponding with the volume of the annular chamber is only needed, aside from leakage. Since, however, oil is continuously supplied, relief bores 40 are provided for discharging the excessive oil. The relief bores 40 connect the annular chamber 31 to the transmission inner chamber 41. It is necessary for a satisfactory pressure build-up in the annular chamber to let the relief bores originate from a point with a peripheral speed which is as low as possible, that is, in the area of the smallest diameter of the annular chamber 31. 
     Pins 42 are secured tightly in the ring gear carrier 19 and extend axially parallel with respect to the axis of rotation of the friction clutch 3. The pins 42 project into corresponding recesses or grooves 43 in the ring piston 22 and prevent same from an undesired rotation relative to the cylinder 21. These recesses 43 have in the illustrated exemplary embodiment an arcuate shape and extend circumferentially concentrically with respect to the axis of rotation curved over an angle β (FIG. 3). The pins 42 rest during normal operation, that is, when the turbine 6 drives, on the right-- referred to FIG. 3--end of the grooves 43. If, for whatever reasons, the speed of the engine 10 increases suddenly to a level where the driven side of the clutch 3, thus the ring piston 22 rotates faster than the driving side and passes same, then this is only possible until the pins 42 contact the left--again referred to FIG. 3--end of the grooves 43 (dashed contour in FIG. 3). The oil supply to the annular chamber 31 between the bores 38 in the wall 21A in the cylinder 21 and the slot 39 in the ring piston 22 is in this position interrupted with the result, that the clutch opens up because of its leakage and interrupts the connection between the turbine 6 and the engine 10. 
     In order to prevent during the relative rotation of the ring piston 22 a sudden change of the pin location from one end of the groove to the other end, the grooves 43 are covered in direction of the annular chamber 31. This can occur with by use of a disk 44 in front of each groove or with one single annular disk covering all grooves 43. The disks 44 (or rather the disk) are (or rather is) pressed against the ring piston 22 by springs 45, so that the oil existing in the grooves 43 will, so to speak, be squeezed out against the resilient disks, thus achieving the desired attenuating effect. 
     The springs 45 are adjusted with respect to their initial force such that they not only press the disks 44 against the ring piston 22, but moreover cause the disks 25, 26, through the action thereon by the ring piston 22, to rest on one another even when the clutch is open. However, the clutch has thereby still so much slippage, that the relative rotation of the ring piston 22 is not hindered. Thus, the springs 45 present only a so-called initial force to make the coupling operation easier. 
     In order to uncouple the turbine in the low speed range of the engine from the engine, in particular in the case of a sudden speed reduction of the engine 10, an interruption of the oil supply to the clutch is provided for such cases. Two pistons 46 are arranged opposite one another in a radial through-bore in the shaft 20, the faces of which rest on one another under the pressure of springs 47 (FIG. 4) and, since the diameter of the pistons 46 is greater than the one of the bore 34, close off the bore 34. Each piston 46 projects thereby up to the axis of rotation of the shaft 20 into the bore 34. Only with an increasing speed of the shaft 20 are the pistons 46 urged radially outwardly under the influence of a centrifugal force against the action or return force of the springs 47 and thus open the bore 34 for the flow of oil. The engaged pistons 46 do not assure a 100% seal. In order to prevent the oil moving through the gap from building up a pressure sufficient for clutch engagement in the annular chamber 31, a second relief bore 48 is provided in the wall 21B of the cylinder 21, which bore 48 connects the annular chamber 31, in the area of its largest diameter, to the transmission chamber 41 and has a very small cross section, that is, substantially smaller than the relief bores 40. Otherwise, it would not be possible, not even during normal operation, for pressure to build up in the annular chamber. 
     The gears 11, 12 in the modification of the exemplary embodiment illustrated in FIG. 2 are also stored in the transmission housing 1. The gear 12 is thereby rotatably supported on a driven shaft 49 and is connected to the driven shaft 49 through the vibration attenuator or damper 4. The gear 13 is mounted on the free end of the driven shaft 49, which free end projects from the transmission housing 1. The gear 13 mates with the gear 14 (not illustrated here). Instead of the gear 13, a flange could also be mounted for connection to the crankshaft 15 or to an additional transmission with the gears 13, 14. The vibration attenuator or damper 4 is in the modification according to FIG. 1 arranged approximately like the gear 11 in FIG. 2 on the shaft 20. The vibration attenuator or damper 4 has in both cases the task of keeping rotational irregularities, which can be very high in particular in the low speed range, substantially away from the clutch and thus also from the turbine 6. In spite of its, if necessary, separate arrangement, the vibration attenuator or damper 4 is an important part of the clutch. This is also true, when the vibration attenuator or damper is arranged at a different, here not mentioned, point between the crankshaft 15 of the engine 10 and the actual clutch 3. 
     The oil, which moves through the relief bores 40, 48 and at other points into the transmission chamber 41, can return through a bore 50 in the transmission housing 1 directly or through not illustrated pipelines to the engine 10.