Classifications

• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
  • F16H—GEARING
  • F16H61/00—Control functions within control units of change-speed- or reversing-gearings for conveying rotary motion ; Control of exclusively fluid gearing, friction gearing, gearings with endless flexible members or other particular types of gearing
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
  • F16H—GEARING
  • F16H61/00—Control functions within control units of change-speed- or reversing-gearings for conveying rotary motion ; Control of exclusively fluid gearing, friction gearing, gearings with endless flexible members or other particular types of gearing
  • F16H61/66—Control functions within control units of change-speed- or reversing-gearings for conveying rotary motion ; Control of exclusively fluid gearing, friction gearing, gearings with endless flexible members or other particular types of gearing specially adapted for continuously variable gearings
  • F16H61/664—Friction gearings
  • F16H61/6649—Friction gearings characterised by the means for controlling the torque transmitting capability of the gearing
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
  • F16H—GEARING
  • F16H15/00—Gearings for conveying rotary motion with variable gear ratio, or for reversing rotary motion, by friction between rotary members
  • F16H15/02—Gearings for conveying rotary motion with variable gear ratio, or for reversing rotary motion, by friction between rotary members without members having orbital motion
  • F16H15/04—Gearings providing a continuous range of gear ratios
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
  • F16H—GEARING
  • F16H15/00—Gearings for conveying rotary motion with variable gear ratio, or for reversing rotary motion, by friction between rotary members
  • F16H15/02—Gearings for conveying rotary motion with variable gear ratio, or for reversing rotary motion, by friction between rotary members without members having orbital motion
  • F16H15/04—Gearings providing a continuous range of gear ratios
  • F16H15/42—Gearings providing a continuous range of gear ratios in which two members co-operate by means of rings or by means of parts of endless flexible members pressed between the first mentioned members
• • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
  • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
  • F16H—GEARING
  • F16H61/00—Control functions within control units of change-speed- or reversing-gearings for conveying rotary motion ; Control of exclusively fluid gearing, friction gearing, gearings with endless flexible members or other particular types of gearing
  • F16H61/66—Control functions within control units of change-speed- or reversing-gearings for conveying rotary motion ; Control of exclusively fluid gearing, friction gearing, gearings with endless flexible members or other particular types of gearing specially adapted for continuously variable gearings

Definitions

• the invention relates to a friction ring transmission with at least two friction wheels arranged at a radial distance and with a friction ring arranged between the friction wheels, which surrounds one of the two friction wheels, the friction wheels and the friction ring being braced by a pressing device, and a method for the Operation of such a friction ring gear.
• Gearboxes of this type fall in the field of continuously variable transmissions (CVT - continuously variable transmission), which also include swashplate gearboxes (toroidal CVT) and chain or belt CVTs, for example from EP 0 466 113, JP 62-258 254, JP 2003-028 257, JP 2001-124 163, JP 06-174 030, JP 06-174 028 or US 3,087,348 or from US 5,184,981, US 5,094,652 or GB 1,600,974 are known.
• CVT continuously variable transmissions
• swashplate gearboxes tilt-toroidal CVT
• chain or belt CVTs for example from EP 0 466 113, JP 62-258 254, JP 2003-028 257, JP 2001-124 163, JP 06-174 030, JP 06-174 028 or US 3,087,348 or from US 5,184,981, US 5,094,652 or GB 1,600,974 are known.
• the friction ring gears differ from these in that one of the gear members surrounds another, or in that the gear member that frictionally connects the two friction wheels has two running surfaces, a first of which only has a first friction wheel and a second one only of a second friction wheel Contact comes. In the case of the other continuously variable transmissions, the contact surface of the variable transmission member alternately comes into contact with the first and the second transmission member.
• both documents disclose a pressing device which, depending on the torque which is transmitted by an output cone of the conical friction ring gear disclosed there, applies a pressing force with which the two cones and the one circulating between the two cones and encompassing the drive cone are encircling Friction ring to be clamped. In this way it can be ensured that a sufficiently high contact pressure is generated at high torques at which there is otherwise a risk of slip.
• EP 0 980 993 A2 discloses a pressing device, the pressing force of which can be regulated or adjusted externally via a hydraulic cylinder.
• the invention proposes a friction ring transmission with at least two friction wheels arranged at a radial distance and with a friction ring arranged between the friction wheels, which surrounds one of the two friction wheels, the friction wheels and the friction ring being clamped via a pressing device, which characterized in that the pressing device comprises at least two partial pressing devices, of which the first of the two partial pressing devices has a shorter reaction time than the second of the two partial pressing devices.
• Such a pressing device comprising two partial pressing devices, for a wide variety of transmissions in which a transmission member, in particular a friction ring or another, frictionally effective transmission member surrounds another transmission member, can be used advantageously.
• the second partial pressing device can be controlled in such a way that the pressing device as a whole provides a pressing force depending on the position of the friction ring - or depending on a similarly significant parameter, such as depending on the transmission ratio - which provides the disproportionate force necessary pressing force, which takes into account a disproportionately high pressing force due to the small radius difference between the friction ring and the cone surrounded by the friction ring.
• the first partial pressing device can still be designed to be relatively simple and therefore not susceptible to faults.
• the reaction time of the first partial pressing device is preferably chosen to be so short that it is possible to react quickly to impacts or the like.
• An arrangement is preferably chosen which is purely mechanical and thus has almost no reaction time. In this way, the pressing device can quickly adapt to short-term fluctuations, as a result of which, in particular, slippage between the gear members rolling on one another can be avoided.
• It may be sufficient, in particular, to directly control the first partial pressing device unregulated and only as a function of critical parameters, such as, in particular, the transmitted torque. In this way, the first partial pressing device - and thus also the entire pressing device - can adjust extremely quickly and reliably to shocks or almost discontinuous or discontinuous changes in the critical parameter.
• the first partial pressure device does not need to be optimized with regard to its characteristic curve, which is dependent on the parameter. Rather, it is important that the first partial pressing device can react appropriately to impacts or discontinuities - in particular with a sufficiently short reaction time.
• An optimal characteristic curve of the entire pressing device is preferably implemented by the second partial pressing device, which can thus preferably be optimized with regard to its characteristic curve or with regard to the characteristic curve of the entire pressing device without being able to react or react to impacts or sudden discontinuities in the short term have to.
• the second partial pressure device is regulated or controlled accordingly, so that the characteristic curve can be selected as best as possible.
• the second partial pressure device can be controlled by different or different characteristics and thus react in detail to the respective requirements.
• the partial pressure device, in particular in its control circuit can be optimized with regard to vibration damping, which generally also leads to a reduction in the response times. However, as already explained above, the latter is not so critical since the first partial contact pressure can react with correspondingly shorter reaction times.
• An arrangement according to the invention can in particular make it possible to considerably minimize the losses in a corresponding transmission.
• the first partial contact pressure device to be optimized in terms of safety or in terms of operational safety, while the characteristic of the second partial contact pressure device is selected such that a safety-related shift in the characteristic curve originating from the first partial contact pressure device is compensated for in a suitable manner.
• a friction ring transmission with at least two friction wheels arranged at a radial distance and with a friction ring arranged between the friction wheels, which surrounds one of the two friction wheels, wherein the friction wheels and the friction ring are clamped via a pressing device
• the pressing device comprises at least two partial pressing devices and in which the first partial pressing device provides a pressing force that is greater than or equal to the pressing force to be provided by the entire pressing device, and the second partial pressing device reduces the pressing force provided by the pressing device.
• the first partial pressing device can provide the necessary pressing force in excess, so that, in particular, short-term fluctuations can be reliably absorbed.
• the second partial pressing device can reduce the excessive pressing force again, so that losses can be minimized without the risk that an insufficient pressing force is available in the event of brief impacts or the like.
• the pressing device in particular the first partial pressing device, can comprise a spring element.
• the spring element can be designed such that it provides a basic load, so that the pressing device with its further mechanical pressing means, which can react variable, only needs to provide the pressing force starting from this basic load.
• the latter applies in particular to the first partial pressing device, so that the characteristic lines to be run through by this partial pressing device can be chosen to be substantially flatter. In particular, this makes it possible to provide linear ramps for mechanical pressing means without having to provide excessive pressing forces which would then have to be compensated for again by the second partial pressing device.
• the second pressing device applies a force opposite to the force applied by the first partial pressing device.
• a reduction in strength can be carried out reliably.
• the first partial pressure device can play its characteristic directly and directly and, if necessary, counteract the reduction in force caused by the second partial pressure device.
• the second partial pressing device preferably partially compensates for the force applied by the first partial pressing device, which, with a suitable design, also leads to the advantages described above, independently of the aforementioned features.
• the contact pressure or the resulting bracing force with the frame or housing is preferably reduced by the second partial pressing device. This leads to a reduction in the total losses, since shocks or rapid changes occur only briefly and therefore only play a minor role over the total operating time.
• a pressing device can be used in a wide variety of gearboxes with gear members rolling on one another. It is particularly suitable for arrangements in which the respective transmission members interact with one another in a frictional or frictional manner or under the risk of slipping if there is insufficient contact force. In particular, such a pressing device can minimize loss in such arrangements.
• a corresponding contact pressure can be applied, for example, by an electromagnetically controlled piston.
• Such an arrangement is small and compact and has a mechanically simple structure.
• the piston can initially close an overflow / refill opening on its stroke. Such an arrangement or process control can ensure at all times that sufficient hydraulic fluid is present between the piston and the pressing device. If a force is applied to the piston, it ensures that the fluid is compressed in the direction of the pressing device until it generates sufficient counter pressure. If the piston is not acted upon, too much fluid can escape through the opening, while on the other hand, if there is too little fluid, fluid can be fed from a reservoir through this opening.
• a gear pump can be provided for the hydraulic control.
• Such a gear pump is relatively inexpensive and also has the advantage that it can also apply variable contact pressures, for example by means of variable rotational speeds or variable torques, in an extremely low-maintenance and reliable manner.
• the gear pump can be driven by an electric motor, wherein a current-dependent torque is preferably provided. This can be done in particular by a current limitation or current regulation, which is generally easier to implement in a motor vehicle than a voltage regulation.
• voltage regulation can be advantageous, particularly in the case of digital control, since it is easier to implement.
• a variable contact pressure can be provided simply and reliably, the gear pump even deliberately not needing to be completely sealed with respect to its wings and can certainly have a slip.
• the necessary contact pressure can be ensured, for example, by a higher speed.
• another pump in particular another pump, which only provides a pressure gradient similar to a gear pump or which has an internal leakage, can also be used.
• a friction ring transmission with at least two operating states in which at least one input member and at least one output member are pressed against one another by means of at least one pressing device with a pressing force that varies depending on the respective operating state, and which is characterized by a pressing device with the one already described above characterized operating condition-contact force characteristic.
• such a varying characteristic curve for the pressing device is advantageous in all friction gears, in which at least one input element and at least one output element engaging around the input element interact with one another in a frictional manner.
• rubbing encompasses any non-positive interaction between two rotating transmission members, whereby non-destructive slip between the two transmission members can preferably occur at high torques.
• this term also includes an interaction that is caused by hydrostatic or hydrodynamic or Electrostatic, electrodynamic or magnetic forces act between the two transmission members, and the present invention thus also includes, in particular, friction gears in which a gap filled with a fluid, such as a gas or a liquid, remains between the actual mechanical transmission members and the speeds, the gap widths , the pressures and the like are dimensioned such that this fluid causes an interaction between the two transmission members, for example due to shear forces.
• this varying characteristic curve f is also suitable r frictional transmission unit, wherein which are provided between the two transmission members, a medium mediating the interaction or a plurality of such media, such as fluids or another transmission member.
• the interaction between the two transmission members is dominated to a relatively large extent by the forces which act on the respective interacting surface of the transmission members.
• the two transmission members can be braced in a suitable manner for this purpose, which can be ensured, for example, by suitable bearings.
• pressure devices can be provided that provide variable contact forces depending on the output torque beyond a certain base load, so that high contact forces can also be generated at high output torques, which increases the transmissible torque of the friction gear accordingly can be. According to the prior art, however, such arrangements lead to relatively high losses in such friction gears, thereby questioning their economy.
• the input member and the output member do not need to be directly connected, rather it is also conceivable that the averaging gear members or the frictional connection averaging measures, such as additional fluids or further interaction mechanisms, may be present. Because of the balance of forces in a transmission, the input link and output link can also be interchanged. However, since such gears are often found in a complex drive train, this differentiation will generally have to be maintained. It goes without saying that the two gear elements can also be pressed against one another by offset degrees of freedom of these gear elements, as long as at least one component of the degrees of freedom used in the pressing or pressing is directed in a suitable manner onto the interacting surface of a corresponding gear element.
• Friction ring gears according to the invention can be operated in different operating states and taking into account different operating state types.
• Such types of operating state can be, for example, input or output torques, speeds, forces or force relationships, pressures or also temperatures, times or the like, and measurement variables proportional to them.
• the respective operating mode types are used in a wide variety of operating conditions, whereby - depending on the specific embodiment or implementation - some operating mode types are only of minor importance or are proportional to other, easily measurable operating mode types.
• a varying characteristic curve can be realized, for example, cumulatively or alternatively with a friction gear, in which the pressing device comprises at least two pressing units.
• the operating state pressing force Characteristic curve can be adapted to desired requirements with relatively simple means. This applies in particular to the various mean slopes of the operating state-contact force characteristic curve, as described above.
• the term “average slope” between two operating states or between an operating state and a rest state describes a value that is determined by an average slope or by an average straight line of the first derivative in the corresponding interval of the operating state-contact pressure characteristic Due to the change in slope, there is the possibility of optimizing the operating state-contact force characteristic curve at least in two respects with regard to the necessities in the drive.Thus, the best possible conditions regarding the driving force depending on the respective specific operating state can be ensured between the two operating states, The contact pressure is optimally selected in relation to the current operating state. This allows losses to be minimized with optimal performance of the friction gear.
• the two pressing units have different operating state-pressing force characteristics as a component of the pressing device.
• the overall characteristic curve of the pressing device can be adapted in a clear and comprehensible manner.
• the two pressing units can each make a first contribution to the pressing force in the first operating state and each make a second contribution to the pressing force in the second operating state, the difference between the first and second contribution of the first pressing device being the difference between the first and second Contribution of the second pressing device differs.
• the respective pressing units make a different contribution to the total pressing force of the pressing device in the respective operating states, as a result of which the characteristic curve of the entire pressing device can be influenced in a structurally simple manner.
• the two pressing units can be designed to act in parallel or in series, independently of the other features of the present invention, with regard to the determination of the operating state and / or pressing force. As a result, as well as by suitable gear ratios with a corresponding coupling The overall characteristic of the pressing device can be easily adapted to the existing requirements.
• the contact pressure units are preferably coupled to one another, wherein the coupling can be designed mechanically or hydrodynamically or hydrostatically. This also applies in particular to the case where the pressure units are each provided separately on a transmission link.
• the coupling can be designed mechanically or hydrodynamically or hydrostatically.
• the first operating state is the lowest torque expected under full load and the second operating state is the highest torque expected under full load. Accordingly, the necessary contact pressure for the lowest torque expected under full load and for the highest torque expected under full load can be determined for a suitable dimensioning of the characteristic curve, so that the corresponding characteristic curve can be formed directly as a straight line between these two points.
• the two pressing units can be varied in terms of their respective pressing force, or in terms of their contribution to the total pressing force of the pressing device, by different operating mode types.
• a contact pressure unit can be varied in terms of its contact force, for example with regard to the input torque or the total load, and a contact unit with regard to the output torque.
• the overall behavior of the friction gear can be adapted to the given requirements in a wide range, so that it can be optimized in particular with regard to its efficiency.
• FIG. 1 shows a first gear according to the invention with a pressing device in a schematic sectional view
• Figure 2 shows the output cone of a second transmission according to the invention with pressing device in a similar representation as Figure 1;
• Figure 3 shows the output cone of a third transmission according to the invention with pressing device in a similar representation as Figure 1;
• Figure 4 is a schematic representation of the force relationships in the embodiments of Figure i;
• Figure 5 is a schematic representation of the force relationships in the embodiments of Figures 2 and 3;
• Figure 6 is a schematic representation of the force relationships in an alternative
• Figure 7 is a schematic representation of the balance of power in an alternative
• Figure 8 is a schematic representation of the balance of power in a further alternative
• FIG. 9 shows a schematic representation of the balance of forces in another exemplary embodiment
• Figure 10 is a schematic sectional view of the alternative indicated in Figure 6 in a similar representation as Figure 1;
• FIG. 11 shows an alternative implementation of the alternative indicated in FIG. 6 in a representation similar to that of FIG. 1;
• Figure 12 is a schematic sectional view of another gear with an alternative pressing device
• FIG. 13 shows a hydraulic control for a transmission according to the invention
• FIG. 14 shows a frictional transmission according to the invention in a schematic sectional illustration
• FIG. 15 shows a schematic section from FIG. 14
• FIG. 16 shows a schematic illustration of the mode of operation of the pressing device from FIGS. 14 and 15;
• FIG. 17 shows the characteristic curve of the inner spherical unit of the arrangement according to FIGS. 14 and 15;
• FIG. 19 shows the characteristic of the entire pressing unit of the arrangement according to FIGS. 14 and 15;
• FIG. 20 shows an alternative characteristic curve of the inner spherical unit of the arrangement according to FIGS. 14 and 15;
• FIG. 21 shows a characteristic curve of the outer ball unit of the arrangement according to FIGS. 14 and 15 adapted to the characteristic curve according to FIG. 20;
• FIG. 22 shows the characteristic curve of the entire pressing unit taking into account the characteristic curves according to FIGS. 20 and 21 of the arrangement according to FIGS. 14 and 15;
• FIG. 23 shows a possible characteristic curve of a pressing device
• FIG. 24 another possible characteristic curve of a pressing device
• FIG. 25 shows a particularly advantageous characteristic curve configuration
• FIG. 26 shows a second friction gear according to the invention in a schematic sectional view
• FIG. 27 the characteristic lines of the input pressure unit of the arrangement according to FIG. 26;
• FIG. 28 shows the characteristic curve of the initial pressing unit of the arrangement according to FIG. 26;
• FIG. 29 the characteristic curve of the entire pressing unit of the arrangement according to FIG. 26;
• FIG. 30 shows a third friction gear according to the invention in a schematic sectional illustration;
• FIG. 31 shows a fourth friction gear according to the invention in a schematic sectional illustration
• FIG. 32 the characteristic lines of the input pressure unit of the arrangements according to FIGS. 30 and 31;
• FIG. 33 the characteristic curve of the initial pressing unit of the arrangements according to FIGS. 30 and 31; and FIG. 34 the characteristic lines of the entire pressing device of the arrangements according to FIGS. 30 and
• the input cone 1 comprises an input cone 1 and an output cone 2, which interact with one another in a manner known per se via an adjustable friction ring 3.
• the input cone 1 is operatively connected to an input shaft 4 and the output cone 2 is connected to an output shaft 5.
• the cones 1, 2 are supported in the radial direction by cylindrical roller bearings 6.
• the cones 1, 2 are braced against each other in the axial direction in this exemplary embodiment by four-point bearings 7A, so that the necessary contact forces can be applied so that torque can be transmitted from the input cone 1 to the output cone 2 and vice versa via the friction ring 3.
• the axial support of the input cone 1 is not explicitly shown in the present figures, but can also be provided, for example, by a four-point bearing 7A or else by an axial cylindrical roller bearing or the like.
• a pressing device 8 is also provided between the output shaft 5 and the output cone 2, while in this exemplary embodiment the input shaft 4 is directly connected to the input cone 1.
• the pressing device 8 is able to vary the axial distance between the output cone 2 and the bearing 7A on the output shaft 5 or - in the tensioned state - to generate correspondingly varying pressing forces.
• bearings 6 and 7A instead of bearings 6 and 7A, other bearing arrangements, such as axial angular contact ball bearings, axial self-aligning ball bearings, axial deep groove ball bearings, tapered roller bearings or similar bearings or bearing types can be combined with one another in order to achieve taper 1, 2 on the one hand to be supported radially and on the other hand sufficiently axially clamped.
• bearings 6 and 7A other bearing arrangements, such as axial angular contact ball bearings, axial self-aligning ball bearings, axial deep groove ball bearings, tapered roller bearings or similar bearings or bearing types can be combined with one another in order to achieve taper 1, 2 on the one hand to be supported radially and on the other hand sufficiently axially clamped.
• hydrodynamic or hydrostatic bearings can be used
• the friction ring 3 can be adjusted in a manner which is not explained in more detail here, but is known, and the transmission ratio of the transmission can be selected in this way. It is understood that the overall arrangement is or is subject to different torques in operation. Since the operative connection between the two cones 1, 2 is a frictional connection, the contact forces should preferably be selected to be sufficiently high so that controllable slip occurs on the friction ring 3. On the other hand, unnecessarily high contact forces would lead to a relatively strong base load, which in turn would affect the efficiency of the friction gear. A manageable and in particular also sufficiently high slip can be advantageous in order to facilitate the control of the transmission, since then only the speed is necessary as a controlled variable, while the torques are correspondingly adapted and transmitted via the contact pressure.
• a torque-dependent contact pressure control is selected in the present exemplary embodiment, but the contact pressure, as will be explained below, can also be selected depending on other operating states.
• the output torque is selected as the manipulated variable for the contact pressure control.
• the pressing device 8 comprises two adjusting disks 9, 10 which have guideways for balls 11 and are supported on the one hand via the adjusting disk 9 on the shaft 5 and on the other hand via the adjusting disk 10 on the output cone 2.
• the adjusting disks 9 and 10 are configured such that the torque is transmitted from the output cone 2 to the adjusting disk 10 via the balls 11 to the adjusting disk 9 and from there to the output shaft 5.
• the guideways for the balls 11 are designed such that an increased torque causes the two adjusting disks 9, 10 to rotate relative to one another, which in turn leads to the balls 11 being displaced along the guideways, as a result of which the adjusting disks 9 and 10 are pressed apart ,
• the torque directly increases the contact pressure due to the sloping guideways.
• the pressing device 8 generates a pressing force that is dependent on the output torque.
• the characteristic curve of the arrangement of the plates 9 and 10 as well as the balls 11 and the spring 12 can only be optimized to a limited extent.
• the Kennlime has areas in which an excessive pressure force is provided.
• the total losses of the corresponding transmission are increased considerably.
• the arrangement from FIG. 1 has force compensation, in particular for partial load ranges.
• this takes place hydraulically, in that a pressure is generated hydraulically between a plate connected to the output shaft 5 and the pressure plate 10, which counteracts the pressure force generated by the balls.
• the excess or unnecessary contact pressure generated by the balls 11 and the spring 12 can be hydraulically compensated by generating a counterforce from a component 13 which is firmly connected to the output shaft 5.
• the compensation takes place in such a way that, depending on the position of the friction ring axially along the two cones, the total contact pressure is controlled accordingly, with a comparatively large contact pressure being provided in the positions in which the friction ring 3 surrounds the strong end of the input cone 1 , which serves to prevent floating of the friction ring 3 reliably.
• the transmission ratio can, for example, also be selected as a parameter proportional to the position. Other parameters that are used accordingly are also conceivable. The further the friction ring 3 is positioned to the pointed end of the input cone 1, the more the contact pressure is compensated for. On the other hand, it is conceivable that the compensation takes place less strongly if it is necessary to master competing effects.
• the corresponding relationships are shown schematically in FIG. 4, the strength of the arrows reflecting the magnitude of the respective forces.
• the hydraulic pressure 14 thus compensates for an excessive force of the balls 11 or the spring 12, so that the bearings 6, 7A are not unnecessarily loaded.
• the arrow 90 indicates external forces from the output shaft 5, the arrow 91 external forces from the output cone and the arrows 92 internal forces.
• the hydraulic pressure 14 is provided via a hydraulic line 15 which is arranged in an additional shaft 16 which is fixedly connected to the shaft 5 via a screw 17.
• the screw 17 also closes a filling opening 18 which, in conjunction with a line 19 and an undercut 20, serves to blow the hydraulic chamber in a reliable manner. free to fill.
• the shaft 16 has a hydraulic seal at its end facing away from the drive shaft 5, so that the hydraulic pressure 14 can be built up or controlled from the outside in the desired manner and without further ado.
• the arrangement according to FIG. 1 also has a mounting body 21, via which the output cone 2 is mounted radially.
• the pressing device 8 can easily be mounted inside the driven cone 2.
• FIG. 2 essentially corresponds to the embodiment according to FIG. 1, so that modules with an identical effect are also provided with identical reference numerals and are not explicitly explained again.
• the base load is not generated by a spring connected in parallel but by a spring 22 connected in series with the pressing device 8, which is supported on the output shaft 5, which in the present embodiment is carried out on a four-point bearing 23, which on the one hand thus the contact pressure between the adjusting disk 9 and the output shaft 5 is transmitted and, on the other hand, the axial bearing of the output cone 2 with respect to the output shaft 5 is used.
• the counter plate 13 which is firmly connected to the output shaft 5 .
• FIG. 2 results in a similar mode of operation as in the embodiment shown in FIGS. 1 and 4.
• a compensating force is generated by the pressure 14, so that the total contact force and thus the bracing force acting on the bearings 6, 7A can be reduced to a minimum via the pressure 14.
• a motor arrangement can also be selected for the second partial pressing device instead of a pressure 14 arrangement, as exemplified in FIG. 3, the exemplary embodiment according to FIG. 3 otherwise corresponding to the exemplary embodiment according to FIG. 2 and as in FIG. 5 appears depicted.
• This arrangement also generates a base load via a spring arrangement 22 connected in series, which is supported on the output shaft 5 via a four-point bearing.
• a threaded bolt 28B is provided in a threaded bore 28A of the output shaft 5, which is supported via a four-point bearing 29 on the positioning plate 10 and on the driven cone 2, with the threaded bore 28A functioning in this arrangement Function of the counter plate 13 corresponds.
• the threaded bolt 28B can be displaced with respect to the shaft 5 via a motor 30, which can be controlled via an electrical line 32 and slip rings 33, and a transmission 31, as a result of which a variable counterforce is generated to the contact force generated by the balls 11 and the spring 22 can be.
• FIG. 6 an arrangement according to the invention can also be implemented without a spring arrangement generating a base load.
• Schematic arrangements which correspond to the relationships according to FIG. 6 are shown in FIGS. 10 and 11.
• a pressing device 8 is provided, in which a positioning plate 9 is supported on the output shaft 5 and has cam tracks for balls 11.
• the ball tracks corresponding to this are provided directly in the output cone 2 instead of in a further positioning plate as in the exemplary embodiments according to FIGS. 1 to 5.
• the second partial pressing device 14 also acts directly on the output cone 2 via a pressure chamber 34.
• the mode of operation corresponds to the mode of operation of the exemplary embodiments already illustrated, so that a detailed discussion is dispensed with.
• the cones 1, 2 are axially supported via axial cylindrical roller bearings 7B.
• the second partial pressing device 14 is primarily controlled as a function of the input torque, which is achieved by means of the input shaft 4, an adjusting disk 35 connected to the input shaft 4, balls 36 and a piston which is non-rotatably connected to the drive cone 1 but is axially displaceable 37 is recorded and passed on hydraulically to the pressure chamber 34 via a line 38.
• the line 38 is in this case sealed through bushings 39 to the assemblies which rotate with the cones 1, 2.
• the second partial pressure device 14 can also be controlled or regulated via a piston 41 as a function of further parameters.
• FIG. 11 represents a mechanical alternative to the embodiment according to FIG. 10, but the input torque determined is transmitted to the second partial pressing device via a lever arrangement 42. Via a servo 43, further manipulated variables can also be used to regulate the second partial pressing device.
• the second partial pressing device or the entire pressing device can be controlled or regulated via various manipulated variables. This can include, in particular, the engine torque, the input speed, the output speed, the adjustment path or the position of the friction ring 3, the temperature of the transmission or a transmission oil, the wheel speeds or, for example, the ABS (anti-lock braking system) signal, an external shock detection or other Parameters.
• the corresponding measured values can be forwarded to the pressing device 8 hydraulically or by motor or in some other way.
• this can be done in particular by pumps, for example gear pumps or by pumps that are already present in a motor vehicle and a corresponding pressure control.
• Piston arrangements and electromotive systems are also conceivable.
• a gear pump 61 driven by an electric motor 62 can be provided, for example, which can require fluid from a reservoir 64.
• a torque 63 can be applied to the gear pump 61 by a voltage 63 applied to the electric motor 62, which rotates it in such a way that the fluid or the pressing device 8 thereby generates a back pressure corresponding to the pressure caused by the torque.
• FIG. 7 A similar mode of operation is shown in FIG. 7, in which the internal forces 92 are provided by means of balls 11 connected in parallel to a hydraulic pressure 14 and a spring arrangement 12 connected in series with them.
• the internal forces 92 are opposed by the external force 90 from the output shaft 5 and the external force 91 by the output cone 2.
• the alternative mode of operation shown in FIG. 8 comprises an arrangement of balls 11 and a hydraulic pressure 14 connected in parallel therewith, the balls 11 and the hydraulic pressure 14 causing internal forces 92. These internal forces 92 are opposed by the external force 90 from the output shaft 5 and the external force 91 by the output cone 2.
• the arrangement according to FIG. 8, like the arrangement according to FIG. 6, does not require an additional spring element.
• the transmission shown in FIG. 12 comprises an input cone 1 and an output cone 2, which interact with one another via an adjustable friction ring 3.
• the input cone 1 is operatively connected to an input shaft 4 and the output cone 2 to an output shaft 5.
• the input cone 1 is supported on the one hand by cylindrical roller bearings and on the other hand by tapered roller bearings 80.
• the tapered roller bearings 80 are particularly well suited to absorbing axially acting forces in addition to radially acting forces.
• the output cone 2 is only supported by cylindrical roller bearings 6, the output shaft 5 of the output cone 2 being additionally supported by means of tapered roller bearings 81.
• the two cones 1 and 2 are braced against each other in the axial direction in such a way that the necessary contact forces can be applied in order to be able to transmit torque via the friction ring 3 from the input cone 1 to the output cone 2 and vice versa.
• a pressing device 8 is also provided between the output shaft 5 and the output cone 2, while in this exemplary embodiment the input shaft 4 is likewise connected directly to the input cone 1.
• the pressing device 8 is also able to vary the axial distance between the output cone 2 and the tapered roller bearing 81 on the output shaft 5 or, in the clamped state, to generate correspondingly varying pressing forces.
• bearings 6, 80 and 81 provided in this exemplary embodiment can also be replaced by other bearing arrangements or combined with other bearing arrangements in order to make the cones 1 and 2 radial on the one hand and sufficient on the other hand to be axially clamped.
• Hydrodynamic or hydrostatic bearings can also be used here.
• the transmission ratio of the transmission illustrated here is selected by moving the friction ring 3, as a result of which different forces, in particular different torques, act on the overall arrangement.
• the pressing device 8 comprises two adjusting disks 9 and 10 which have guideways for balls 11.
• the adjusting disc 9 or 10 are configured such that the torque is transmitted from the output cone 2 to the adjusting disc 10 via the balls 11 to the adjusting disc 9 and from there to the output shaft 5.
• the guideways for the balls 11 are configured in such a way that an increased torque causes the two adjusting disks 9 and 10 to rotate relative to one another, which in turn leads to the balls 11 being displaced along the guideway, as a result of which the adjusting disks 9 and 10 are pressed apart , Ideally, rotational movements between the two adjusting disks 9 and 10 are not carried out if the arrangement is essentially rigid.
• the torque is due to the inclined guideways immediately an increase in contact pressure.
• the pressing device 8 generates a pressing force that is dependent on the output torque.
• the arrangement described here as a mechanical device, advantageously has extremely short reaction times and can react very well, in particular, to impacts in the drive train on the output side.
• the adjusting disks 9 and 10 are pressed apart by means of a spring arrangement 12, which provides a certain basic load in the pressing device 8. Since the characteristic of the present pressing device 8 can only be optimized to a limited extent, the pressing device 8 has a force compensation, in particular for partial load ranges. In this exemplary embodiment, this is done hydraulically by hydraulically generating a pressure between a plate of the adjusting disk 10 connected to the output shaft 5, which pressure counteracts the contact pressure generated by the balls 11 and the springs 12. In this way, the excess or unnecessary contact pressure generated by the balls 11 and the springs 12 can be hydraulically compensated.
• the pressure is provided via a hydraulic line 15, which is arranged in an additional shaft 16.
• An oil chamber 82 is provided between the pressing device 8 and the output cone 2. Centrifugal forces, which act in particular on the oil in the pressing device 8, are better compensated for by the oil arranged in this oil chamber 82.
• a reservoir 64 is provided in order to have a sufficiently large amount of oil available to regulate the pressing device 8.
• a torque 63 can be applied to a pump 61 by means of a voltage 63 applied to the electric motor 62, by means of which the pump 61 is adjusted in such a way that the fluid or the adaptation device 8 thereby generates a back pressure corresponding to the pressure caused by the torque.
• FIG. 13 A suitable alternative is the example shown in FIG. 13, in which a coil 46 is provided on a housing 44 via a spacer 45, within which a core 47 with a piston 48 is arranged, which by means of a spring 49 in the housing 44 is pressed. If a current is applied to the coil 46, the core 47 is pressed into the center of the coil 46 against the spring force 49, so that the piston 48 pushes into a cylinder 50 and in this way in this cylinder 50 and in a line connected to it 51 generates a variable pressure as a function of the voltage applied to the coil 46.
• the line 51 can be connected, for example, to the feed 26 from the exemplary embodiments according to FIGS. 1 and 2 or to the line 38 from the exemplary embodiment according to FIG.
• An opening 52 is provided in the cylinder 50, which is sealed first when the piston 48 moves forward.
• This opening 52 is connected to an overflow / refill container 53, so that hydraulic fluid can be refilled or filled in the relaxed state of the overall arrangement in order, for example, to counter leakage or excess pressure caused by external influences. genzu mention. It goes without saying that such an electrical control of a hydraulic piston and / or such a leakage protection can also advantageously be used independently of the other features of the present invention.
• the friction gear shown in FIGS. 14 to 22 and explained including its characteristic curves has an input cone 101 and an output cone 102 which interact with one another via an adjustable friction ring 103.
• the input cone 101 is operatively connected to an input shaft 104 and the output cone 102 is connected to an output shaft 105.
• the cones 101, 102 are supported in the radial direction by cylindrical roller bearings 106 (only shown schematically in FIG. 14).
• the cones 101, 102 are clamped in the axial direction in this exemplary embodiment by axial cylindrical roller bearings 107, so that the necessary contact forces can be applied so that torque can be transmitted from the input cone 101 to the output cone 102 and vice versa via the friction ring 103 ,
• a pressing device 108 is also provided between the output shaft 105 and the output cone 102, while in this exemplary embodiment the input shaft 104 is connected directly to the input cone 101.
• the pressing device 108 is able to vary the axial distance between the output cone 102 and the axial cylindrical roller bearing 107 on the output shaft 105 or - in the tensioned state - to generate correspondingly varying pressing forces due to a spring arrangement 109.
• bearings 106 and 107 can be combined with one another in order to produce cones 101, 102 on the one hand to be supported radially and on the other hand sufficiently axially clamped.
• bearings 106 and 107 other bearing arrangements, such as axial angular contact ball bearings, axial self-aligning spherical bearings, axial deep groove ball bearings, tapered roller bearings or similar bearings or bearing types can be combined with one another in order to produce cones 101, 102 on the one hand to be supported radially and on the other hand sufficiently axially clamped.
• hydrodynamic or hydrostatic bearings can be used, for example.
• the friction ring 103 can be adjusted in a manner which is not explained in more detail here, but is known, and the transmission ratio of the transmission can be selected in this way. It is understood that the overall arrangement is subject to different torques in particular during operation. Since the operative connection between the two cones 101, 102 is a frictional connection, the contact forces should preferably be selected such that controllable slip occurs on the friction ring 103. On the other hand, unnecessarily high contact forces would lead to a relatively strong base load, which in turn would affect the efficiency of the friction gear. For this reason, a torque-dependent contact pressure control has been selected in the present exemplary embodiment, but the contact pressure can also be selected depending on other operating states. As can be seen directly from FIGS. 14 and 15, the output torque is selected as the manipulated variable for the contact pressure control, and also other types of operating states, such as the total load or the input torque, can be used in this regard, as will be illustrated on the basis of the exemplary embodiments explained below.
• the pressing device 108 comprises two pressing units 110 and 111 connected in parallel with respect to their torque measurement and in series with respect to their pressing force action, which are each represented by inner balls 112 and outer balls 113 (see FIG. 15).
• the balls 112, 113 each run in ball tracks, which in the tapered or shaft-side pressure plates
• the shaft-side pressure plates 114 and 115 are arranged in a rotationally fixed manner with respect to the output shaft 105, while the cone-side pressure plate 116 is arranged in a rotationally fixed manner with respect to the output cone 102.
• the pressure plates 114, 115, 116 are provided.
• FIGS. 1-10 From there via the balls 112, 113 and via the pressure plate 115 and the bearing 118 to the pressure plate 114 and from the pressure plate 114 via the bearing 117 to the output shaft 105, the pressure plates 114, 115, 116 axially against the spring force of the spring assemblies 109 and against a pressure bearing 120, which is supported by an axial cylindrical roller bearing 121 and a bearing plate 122 on the driven cone 102, and in this way generate a torque-dependent pressure force depending on the curve tracks.
• FIG. 16 shows schematically in flat form the interaction of the two pressing units 110 and 111, whereby identical reference numbers are also assigned to modules which have the same effect as the modules from FIGS. 14 and 15.
• the balls 112, 113 run in configured ball tracks with different inclinations ⁇ and ⁇ . More complex tracks can also be used, if necessary, linear tracks in particular being advantageous for reasons of reliability, for example against play or thermal effects.
• Given a displacement or given torque as is shown, for example, in the lower part of FIG. 16 using an adjustment path V compared to the arrangement of the upper part of FIG. 16, these ball tracks each require a stroke H1 or H2, which results in a total stroke G.
• the stroke H1 is limited by the stop, so that the total stroke G does not depend linearly on the adjustment path V.
• the ball tracks can, for example, be designed in such a way that the characteristic curves shown in FIGS. 17 and 18 result. Because of the torque-related parallel connection, the characteristic curve shown in FIG. 19 follows from this, the torques adding up because of the parallel connection with regard to the torque and because of the series connection with respect to the axial contact force, the contact pressure is identical for both contact units. When the shoulder 123 is reached, only the outer pressing unit 111 contributes with its characteristic curve to the overall characteristic curve.
• FIGS. 20 to 21 show another characteristic curve configuration, the particularly positive overall characteristic curve (FIG. 22) resulting from the negative slope in the inner pressing unit.
• the pressing units in the present exemplary embodiments have an operating state-pressing force characteristic curve or a torque-pressing force characteristic curve with a substantially constant slope.
• a characteristic curve adapted to the respective requirements can be achieved despite these essentially constant gradients. This is possible, inter alia, in that the two pressing units 110, 111 each make a first contribution to the pressing force with a first torque and make a second contribution to the pressing force with a second torque, the difference between the first and second contribution of the first Pressing device 110 deviates from the difference between the first and second contribution of the second pressing device 111.
• the cam track can be countered, for example, by giving the cam track a variable slope, as shown in FIG. 24.
• the characteristic curve has an essentially constant slope in the operating range between 50 Nm and 350 Nm and drops below the operating range to a contact pressure in the vicinity of 0 N, in particular below 1 N, in the idle state (0 Nm).
• the base load in the overall system is significantly reduced, which can increase the overall efficiency.
• a variable slope of the cam track in a pressing unit hides ranz problems in itself, which the present invention prevents by using at least two pressing units, as already described above.
• the invention proposes that, as shown in particular in FIGS. 24 and 25, the operating state-contact force characteristic curve in an operating range (cf. 50 Nm to 350 Nm in FIGS. 24 and 25) has a lower mean gradient than below this operating area.
• an operating range cf. 50 Nm to 350 Nm in FIGS. 24 and 25
• arrangements are also conceivable which make a characteristic curve similar to the characteristic line shown in FIG. 19 with an operating range between 100 Nm and 350 Nm appear desirable.
• Such a characteristic lime can also be realized in particular by two pressing units with a low tolerance sensitivity.
• FIG. 26 This arrangement essentially corresponds to the arrangement shown in FIGS. 28 and 29, the cones 101 and 102 in this arrangement, apart from being supported by the cylindrical roller bearings 106, being supported in the axial direction by angular contact ball bearings 124.
• the pressing device is formed by two pressing units 125, 126. 28 and 29, a pressing unit 125 is arranged on the output cone 102 and the other pressing unit 126 is arranged on the input cone 101.
• the pressing units 125, 126 have the characteristic curves shown in FIGS. 27 and 28. This results in the characteristic lime shown in FIG. 29, which essentially corresponds to the characteristic lime of the output pressing unit 125, but changes to a horizontal at low moments depending on the load.
• the slope of the characteristic curve of the output contact pressure unit 125 is selected such that this characteristic curve intersects the ideal full-load characteristic curve in the operating interval, so that a sufficiently high contact force results at high output torques.
• the overall arrangement is designed in such a way that the ideal full-load characteristic is not undershot even at full speed in the lower speed range. In the case of partial loads, the ideal full-load characteristic curve can be fallen below depending on the load, so that the total load in the system is further reduced, even though excessive contact forces are provided in full-load operation.
• the slope of the characteristic curve for the output pressure unit 125 By choosing the slope of the characteristic curve for the output pressure unit 125, its intersection with the ideal full-load characteristic curve can be shifted in order to minimize the total losses in this way.
• the slope of the characteristic curve of the Output pressure unit 125 cannot be selected equal to the slope of the ideal full-load characteristic curve in the operating range, since then the effects from the second pressure unit 126 do not come into play.
• the pressing units 125, 126 are each arranged in different transmission members of the friction transmission, as is already the case with the embodiment according to FIG. 26.
• the pressure units 125, 126 each comprise ball arrangements 127, 128, which are each supported on pressure plates 129, 130 of the input shaft 104 and the output shaft 105, respectively.
• the balls 128 are supported on a pressure plate 131 which is axially displaceable but non-rotatably with respect to the input cone 101.
• This pressure plate also serves as a piston for hydraulic feedback 132 with a piston 133, which in turn is connected to the pressure plate 130.
• a further pressure plate is not provided in the pressure unit 125 on the output side, since the balls 127 are otherwise arranged directly on the output cone 102, a separate pressure plate for receiving the corresponding cam tracks also being able to be provided in this regard.
• the hydraulic feedback 32 is passed through bushings 134, 135 into the interior of the cones 101, 102, and instead of such a hydraulic feedback 132, a mechanical system 135 corresponding to the arrangement according to FIG. 31 can also be provided, which is equipped with corresponding plates 136 , 137 of the pressing units 125, 126 interacts.

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• Engineering & Computer Science (AREA)
• General Engineering & Computer Science (AREA)
• Mechanical Engineering (AREA)
• Friction Gearing (AREA)
• Automatic Assembly (AREA)

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• 2004
  • 2004-06-17 JP JP2006515679A patent/JP2006527821A/ja active Pending
  • 2004-06-17 BR BRPI0411486-8A patent/BRPI0411486A/pt active Search and Examination
  • 2004-06-17 DE DE112004001100T patent/DE112004001100D2/de not_active Ceased
  • 2004-06-17 KR KR1020057024293A patent/KR101171123B1/ko active IP Right Grant
  • 2004-06-17 WO PCT/DE2004/001231 patent/WO2004113766A1/de active Application Filing

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