Improved switching of axial couplings with flat teeth under load

Abstract

The present invention relates to an axial coupling having a first shaft and a second shaft, which can be coupled and decoupled controlled by an actuator, comprising a first active surface pairing with a first helix angle and a second active surface pairing with a second helix angle, wherein the first helix angle comprises an angular range from 0 degrees to 30 degrees relative to an imaginary plane through a rotation axis and the second helix angle comprises an angular range from 0 degrees to 30 degrees relative to an imaginary plane through a rotation axis, and the sum of the helix angles via which a torque-dependent axial force can be generated and which assists the decoupling process of the axial coupling under load comprises an angular range from 3 degrees to 35 degrees. The present invention further relates to a method for dimensioning an axial coupling.

Description

[0001] Improved switching of axial couplings with flat gears under load

[0002] 1. Technical area

[0003] The present invention relates to an arrangement of planetary gears or spur gears which can be switched into different gears and which has at least one axial clutch, which is optionally held closed by a pressure force of a spring on an axially toothed clutch ring which engages with control fingers in a control groove of a control cylinder, by the rotation of which for the purpose of a gear shift the clutch ring engages or disengages the clutch in question by a switching force occurring on the groove side wall which is directed against the pressure force.

[0004] 2. State of the art

[0005] Such transmissions with axial clutches for the purpose of setting different gear stages are previously known, for example, as planetary gears from German patent application DE 10 2018 007 326 A1 and described in German patent application DE 10 2021 129423 A1. All gears in these transmissions can be shifted even under operating load. However, during individual shift transitions, considerable shifting forces are required to disengage a clutch against the load-related restraining forces that occur between the clutch elements, so that a reduction in the drive torque of the crank drive and / or the auxiliary motor is indicated during the corresponding shifting operations. DE 10 2021 129 423 A1 describes a rear-mounted transmission having helical gear teeth used for the indirect shifting of the sun gear clutch KSR. The shifting operation is triggered by the change in state of the clutch KHR.The sun gear clutch does not have an actuator, which is controlled via a control groove in a formwork drum. Furthermore, after the ring gear clutch KHR is engaged, the KSR clutch is in freewheeling mode because the SR sun gear rotates faster than the carrier.

[0006] 3. Task

[0007] The object of the present invention is to improve the switchability of axial clutches of planetary gears or spur gears so that they can be safely coupled even under drive load and uncoupled with low switching forces by an actuator.

[0008] 4. Summary of the invention

[0009] The problem is solved according to the features of independent claim 1 and independent claim 14. Further advantageous embodiments of the solution can be found in the respective dependent claims.

[0010] The solution consists in the toothing, in particular a) a coupling toothing and a plug-in toothing or b) a coupling toothing and a running toothing of the spur gears or c) a plug-in toothing and a running toothing of the spur gears of a planetary gear or a spur gear, is / are designed with regard to the inclination of the toothing in such a way that the torque-dependent axial forces occurring under drive load act in the opening direction of an axial coupling and largely compensate for the static friction acting counter to the decoupling of the axial coupling between the contacting coupling teeth and the contacting plug-in connection teeth or between the contacting coupling teeth and the contacting impeller teeth, or between the contacting plug-in connection teeth and the contacting impeller teeth,The resulting reaction forces occurring on the clutch elements are absorbed in support elements on adjacent transmission components.

[0011] The spur gear or planetary gear is arranged in a housing which is connected, either fixedly or by means of a torque arm which is connected to an axle in a rotationally and axially fixed manner, to the frame of a vehicle which is driven by at least one wheel and is movable over a roadway. The frame establishes a directional reference to the roadway. Typically, at least two wheels establish the directional reference between the frame and the roadway. In the case of only one drive wheel without any other wheels, the directional reference can be established by the rider (unicycle) or by a control system (segway). In the present case, it is preferably assumed that the vehicle is driven by at least one of the wheels.

[0012] The transmission is arranged between two shafts. The drive can, for example, comprise a shaft driven by muscle power via a pedal and / or a shaft driven by a motor.

[0013] A shaft in the sense of this invention is a component that can absorb or transmit torque.

[0014] At least one axial coupling is arranged within the transmission, which couples an incoming shaft to an outgoing shaft in such a way that coupling or decoupling takes place between the incoming shaft and the outgoing shaft in a controlled manner via an actuator.

[0015] The input and output shafts are rotatably mounted in the gear housing of the spur gear unit and, as a rule, in the gear housing of the planetary gear unit. In the case of the planetary gear unit, one of the shafts can be connected to the gear housing or the main shaft in a rotationally fixed manner.

[0016] To improve the switching behavior, torque-dependent axial forces are applied by a) a clutch toothing with a helix angle β K surface normals of the active surfaces tilted towards the rotation axis, so that the angle between the active surface and the rear flank of the coupling teeth is increased or b) a helical running gear with a helix angle β L or c) a helical spline with a helix angle β S generated.

[0017] If the respective helix angle β K or β L or β S is equal to the arctangent of the friction coefficient μ of the opposing pairs of effective surfaces in the respective material pairing, the axial force corresponds in magnitude to the axial component of the friction force. This angle is referred to below as the friction angle ρ.

[0018] The friction angle ρ of the active surface pairs depends on the material pairing of the opposing coupling materials in the axial coupling and the lubrication between the active surface pairs. In principle, the temperature at the axial coupling also has an influence, but can be neglected within the temperature range of typical gearbox applications. Another factor is the time the active surface pairs are under load. The active surface pair refers to the surfaces that transfer the forces caused by the applied torque from the driving clutch part to the output clutch part when the clutch is engaged.

[0019] If the axially movable part of an axial coupling is given the degree of freedom necessary for engagement with the axially fixed coupling part by means of a spline, whereby the axial mobility of the component carrying the coupling teeth is achieved by the inherent axial displacement of the component with the coupling teeth relative to an axially closely guided, driving component, the axial coupling can be opened and closed using the provided actuator. This enables potential rotation of the coupled shaft through the flat-toothed axial coupling.The axial force required to open the axial coupling to overcome the friction between the effective surfaces of the spline and the effective surfaces of the flat-toothed coupling teeth depends, in addition to the torque to be transmitted, on the material pairing of the effective surface pair of the spline, their pitch circle diameter, and the material pairing of the effective surface pair, as well as the mean pitch circle diameter of the coupling teeth. The mean pitch circle radius of the coupling teeth is the mean radius of the circular ring that just completely encloses the radially aligned coupling teeth.

[0020] If the axially movable part of an axial coupling is given the degree of freedom necessary for engagement with the axially fixed coupling part by means of a running gear, whereby the axial mobility of the gear carrying the coupling teeth is achieved by the inherent axial displacement of the gear with the coupling teeth relative to a closely guided, axially guided driving gear, the axial coupling can be opened and closed using the provided actuator. This enables potential rotation of the coupled shaft through the flat-toothed axial coupling.The axial force required to open the axial coupling to overcome the friction between the effective surfaces of the running gear teeth and the effective surfaces of the flat-toothed coupling teeth depends, in addition to the torque to be transmitted, on the material pairing of the effective surface pair of the running gear teeth, their pitch circle diameter and on the material pairing of the effective surface pair and the average pitch circle diameter of the coupling teeth.

[0021] If the axially movable part of an axial coupling is given the degree of freedom required for engagement with the axially fixed coupling part by means of a running gear and a spline, whereby the axial mobility of the gear carrying the coupling teeth is achieved by the inherent axial displacement of the gear with the clutch gear and the spline relative to an axially closely guided, driving gear, the axial coupling can be opened indirectly via an actuator when opening into freewheel mode and closed by means of a return spring when the drive torque of the driving gear is lost. This enables a potential rotation of the disengaged shaft relative to the previously coupled shaft.

[0022] The axial force required to open the axial coupling to overcome the friction between the active surfaces of the running gear and the active surfaces of the spline depends, in addition to the torque to be transmitted, on the material pairing of the active surface pair of the running gear, their pitch circle diameter, and the material pairing of the active surface pair and the pitch circle diameter of the spline. Since the axial coupling switches to freewheel mode upon opening, no frictional forces arise between the active surfaces of the coupling gear. With a coupling of this type, the drive shaft is changed under load. The following applies: the smaller the respective radius, the greater the friction force for a given torque.

[0023] In the variants according to the invention, there are two pairs of active surfaces at which a first and a second friction force are generated, the friction values ​​of which depend on a first and a second material pairing.

[0024] When opening the axial coupling, both frictional forces must be overcome. The respective effective pitch circle radii are generally different, so the different radii are also taken into account in the invention.

[0025] In the following, some reference symbols are inserted for better understanding, which are also used in the detailed description of the figures and the derivation of the calculation method.

[0026] The present invention further achieves the object by means of a method for dimensioning the axial couplings according to the invention, taking into account the geometric and tribological conditions, with the following steps:

[0027] • Determination of the friction coefficient μ1between the first

[0028] Material pairing MP1 involved active surface pair WP1 and the friction coefficient μ2between the material paired with a second

[0029] Material pairing MP2 involved active surface pair WP2 using standard methods on the object and / or from a table with

[0030] Material properties.

[0031] • Determination of the effective pitch circle radius r1 of the first effective surface pairing WP1 and the effective pitch circle radius r2 of the second effective surface pairing WP2 from the design data.

[0032] • Determination or specification of a first helix angle β1 of the first effective surface pairing WP1 and a second helix angle β2 of the second effective surface pairing WP2 from design data that are load-independent, namely the friction coefficients of the first material pairing MP1 and the second material pairing MP2 μ1 and μ2, the pitch circle radii of the first effective surface pairing WP1 and the second effective surface pairing WP2 r1 and r2, whereby, for example, in the case of complete compensation of the friction forces, the angle β2 can be determined as a function of β1: β2 = 2 * arctan ( [ a * b + (( a 2 + 1 )*(b2 + 1 )) 1 / 2 - 1 ] / [ a + b ] ) with: a = μ2b = - r2 / r1* [ sin(β1) - μ1*cos(β1) ] / [ cos(β1) + μ1*sin(β1) ]

[0033] A sufficiently good approximation for angles up to 30° is given by: β2= ρ2- (r2 / r1) * (β1- ρ1) with: ρ1= arctan(μ i ), i = 1 ,2

[0034] A simple solution is to compensate the friction forces individually, ie: β2= ρ2and β1= ρ1.

[0035] In a first preferred and alternative embodiment of the axial coupling, to support the decoupling process of the axial coupling, it comprises a spline in a first active surface pairing WP1 with material pairing MP1 and a coupling spline in a second active surface pairing WP2 with material pairing MP2. In this embodiment, the helix angle β1 is replaced by a helix angle β S of the spline. The helix angle β2 is determined by a helix angle β Kthe clutch teeth.

[0036] The pitch circle radius r1 is formed by a pitch circle radius r S of the spline. The pitch circle radius r2 is formed by a mean pitch circle radius r K the clutch teeth.

[0037] The friction coefficient μ1 results from the first material pairing MP1 of the spline with a friction coefficient μ S . The friction coefficient μ2 results from the second material pairing MP2 of the coupling teeth with a friction coefficient μ K .

[0038] In a further alternative embodiment of the axial coupling, to support the decoupling process of the axial coupling, it comprises a plug-in toothing in the first active surface pairing WP1 with material pairing MP1 and a running toothing in the second active surface pairing WP2 with material pairing MP2.

[0039] In this embodiment, the helix angle β1 is replaced by a helix angle β S of the spline. The helix angle β2 is determined by a helix angle β L the running gear teeth.

[0040] The pitch circle radius r1 is formed by a pitch circle radius r S of the spline. The pitch circle radius r2 is formed by a pitch circle radius r L the running gearing.

[0041] The friction coefficient μ1 results from the first material pairing MP1 of the spline with a friction coefficient μ S . The friction coefficient μ2 results from the second material pairing MP2 of the running gear with a friction coefficient μ LIn a further alternative embodiment of the axial coupling, to support the decoupling process of the axial coupling, the latter comprises a coupling toothing in the first active surface pairing WP1 with material pairing MP1 and a running toothing in the second active surface pairing WP2 with material pairing MP2.

[0042] In this embodiment, the helix angle β1 is replaced by a helix angle β K of the coupling teeth. The helix angle β2 is determined by a helix angle β L the running gear teeth.

[0043] The pitch circle radius r1 is formed by a mean pitch circle radius r K of the coupling teeth. The pitch circle radius r2 is formed by a pitch circle radius r L the running gearing.

[0044] The friction coefficient μ1 results from the first material pairing MP1 of the coupling teeth with a friction coefficient μ K. The friction coefficient μ2 results from the second material pairing MP2 of the running gear with a friction coefficient μ L .

[0045] The schematic drawings 1b - 1e show the relationships described below.

[0046] With the axial coupling 100, a first shaft 200 can be coupled and uncoupled with a second shaft 220, controlled by an actuator 140, comprising a first active surface pairing WP1 with a first helix angle β1 and a second active surface pairing WP2 with a second helix angle β2, wherein the first helix angle β1 comprises an angular range of 0 degrees to 30 degrees relative to an imaginary plane through a rotation axis 210 and the second helix angle β2 comprises an angular range of 0 degrees to 30 degrees relative to an imaginary plane through a rotation axis 210 and the sum of the helix angles β1 plus β2, over which a torque-dependent axial force A GEScan be generated and which supports the decoupling process of the axial coupling 100 under load, covers an angular range of 3 degrees to 35 degrees.

[0047] In a first embodiment, the axial coupling 100 comprises in the first active surface pairing WP1 a spline 113 with at least one tooth 112 and the first helix angle β1 comprises a helix angle βs and in the second active surface pairing WP2 a coupling spline 114 and the second helix angle β2 comprises a helix angle β K .

[0048] In this case, a coupling ring 130 of the coupling toothing 114 can be provided, which comprises an asymmetric toothing 131 with at least one tooth 112, further comprising a flank 132, which has the angle β K has.

[0049] The coupling ring 130 is designed to be controllable in an opening direction x by an actuator 140, wherein the axial force A GESthe opening of the axial coupling 100 in opening direction x, is supported.

[0050] In a further embodiment, the axial coupling 100 comprises in the first active surface pairing WP1 a spline 113 with at least one tooth 112 and the first helix angle β1 comprises a helix angle βs and in the second active surface pairing WP2 a running spline 600 and the second helix angle β2 comprises a helix angle β L .

[0051] In a further embodiment, the axial coupling 100 comprises in the first active surface pairing WP1 a coupling toothing 114 with at least one tooth 112 and the first helix angle β1 comprises a helix angle β K and the second effective surface pairing WP2 comprises a running toothing 600 and the second helix angle β2 comprises a helix angle β L. The first material pairing MP1 and the second material pairing MP2 of the axial coupling 100 can each comprise one of the material pairings of steel, bronze, beryllium copper, aluminum, aluminum multi-material bronze, ceramic, carbon or carbon fibers individually, for example steel on steel, aluminum on aluminum, etc., or in combination, for example steel on aluminum or steel on bronze, etc., in order to ensure safe coupling and uncoupling.

[0052] The following examples explain preferred embodiments in detail. It is taken into account that the friction coefficients depend on the material pairing, heat treatment such as hardening, lubrication, contact duration, etc. This means that a considerable variation in the friction values ​​is taken into account when designing the solution for each application.

[0053] German patent application DE 10 2021 129 423 A1 describes a ten-speed hub gear consisting of a five-speed input gear EGG and a secondary gear NSG, which includes a direct gear and a secondary gear with an overdrive speed ratio. In direct gear, the secondary gear NSG rotates in a block with the KHR clutch disengaged, and in overdrive, the KHR clutch is engaged. Due to the high gear ratio, approximately 35% of the input torque is transferred to the hub sleeve, leaving 65% for the countertorque to be absorbed by the main axle. This torque is shifted by the KHR clutch when the secondary gear NSG is shifted from overdrive to direct gear. The axially fixed part of the KHR clutch is firmly connected to the ring gear HR5 via a side wall (see Fig. 3a).The axial limitation of the movement range of the KHR clutch is ensured by the center sleeve MH and the output-side flange of the main axle HA. The axially movable part of the KHR clutch is directly connected to the main axle HA via the spline and is controlled by a shift finger from the control groove NHR in the shift drum ST. To disengage the KHR clutch, both the static friction between the movable clutch part and the main axle HA and the static friction between the contacting surfaces of the clutch splines must be overcome.

[0054] In the following, embodiments of the invention are described as it is used within an overall device such as a coupling assembly or a system.

[0055] Preferably, the torque-dependent axial forces in the opening direction of the axial clutch are generated by a rotated position of the normals of the active surfaces of the meshing clutch teeth. The normals advantageously form an angle β with the orthogonal to the rotational axis of the transmission. K that in the limiting case, a compensation of the frictional forces against the axial forces is achieved, according to equation / 10 / . However, if only partial compensation is achieved by a smaller angular position of the normal, the control groove NHR is retained, as shown in Fig. 3b. In the case of full compensation or overcompensation by the helical gearing(s), the control groove NHR is designed with a constant width adapted to the shift fingers over the entire angle of rotation of the shift drum, as shown in Fig. 3c, so that the axial clutch remains securely closed in the engaged state.

[0056] Another preferred embodiment of the present invention is shown in Fig. 4d. A load-switchable axial clutch relates to a block-type circulating clutch that connects or disconnects two rotating parts in a switching manner. This is, for example, a K70 clutch of a ten-speed hub transmission from German patent application DE 10 2021 129 423 A1. The reference numerals used there are also used and supplemented here.

[0057] The K70 coupling connects a sun gear SR3 to a center sleeve MH, which also serves as a partial web for the output-side part of an input gear EGG. The movable part of the axial coupling is guided axially and rotationally secured inside the center sleeve MH by means of a spline. The axially fixed part of the K70 coupling is housed in one piece on the output side of the axially closely guided sun gear SR3. The movable part of the K70 coupling has an annular collar inside, which serves as a contact surface for a sliding ring guided axially and rotationally secured on the main axis HA.

[0058] The sliding ring is designed so that the collar inside the movable clutch part K70 rotates freely, tightly guided by a rotationally secured, spring-loaded stop disk and an axially offset stop surface of the sliding ring. The axial length of the shoulder is only slightly larger than the wall thickness of the inner annular collar of the movable clutch part. This prevents a clamping effect. A spring F71 acting in the opening direction secures the open position of the clutch K70. A spring F70 acting in the closing direction supports the closing process of the clutch K70, as a track change occurs in a control groove N70 on a shift drum ST.

[0059] If the static friction is only partially compensated, the sliding ring and thrust washer AS are installed rotated 180 degrees, and spring F71 is omitted. This is because, supported by the axial forces, clutch K70 is now opened against the spring force of spring F70. This also applies to the other examples.

[0060] With the K70 coupling, a helical spline and a helical coupling spline can be combined. The helical coupling spline is preferably used.

[0061] A further preferred embodiment of the present invention of a load-switchable axial clutch relates to a clutch that sits on the outer surface of a rotating sleeve arranged coaxially with a main axis HA and is actuated through this sleeve. An example of this embodiment is clutch 90 from German patent application DE10 2018 007 326 A1. In contrast to the prior art, the control of the axial clutch, which is assisted by helix angles during decoupling, is advantageously carried out selectively via at least one actuator. The compensation of the frictional forces can preferably be distributed between the spline and / or the clutch spline.

[0062] A further preferred embodiment of the present invention of an axial clutch that can be switched under load relates to an axial clutch that works as a follow-on lock for a sun gear and must be switched under load. Clutches of this type can be found in German patent application DE 10 2018 007 326 A1 and in German patent application DE 10 2021 129 423 A1, in particular clutches 40 and K40 of the input gearbox EGG therein. This input gearbox has a complete planetary gear set with sun gear, planet gear, and ring gear in the first sub-gearbox, the input of which is the planet carrier and the output is via the ring gear. To change gears, the sun gear is set or released in a rotationally fixed manner. This requires a follow-on lock. When releasing the axially closely guided sun gear, the axial clutch must be opened under load.

[0063] According to the invention, within the sun gear in this exemplary clutch assembly or this exemplary system or in this exemplary overall device, the movable clutch part is guided concentrically to the latter by means of a spline in a rotationally fixed and axially displaceable manner, and to support the switching process, the spline is preferably provided with a helix angle ßs and the effective surfaces of the clutch teeth are arranged relative to an imaginary plane through the rotation axis 210 by the angle β Ktwisted. The non-rotatable and axially fixed part of the axial coupling is preferably connected to the main axis HA via a spline. The axial locking is achieved, as shown in Fig. 5a, with screws that engage in holes adapted to the screw. Fig. 5c shows the control of the axial coupling for the case of compensation and overcompensation of friction forces. The control of the axial coupling corresponds to that of the K70 coupling. In the case of undercompensation, the description there also applies.

[0064] The present invention comprises an axial coupling with a first shaft and with a second shaft, which can be coupled and uncoupled under the control of an actuator, comprising a first active surface pairing WP1 with a first helix angle β1 and a second active surface pairing WP2 with a second helix angle β2, wherein the first helix angle β1 comprises an angular range from 0 degrees to 30 degrees and the second helix angle β2 comprises an angular range from 0 degrees to 30 degrees and the sum of the helix angles β1 plus β2, over which a torque-dependent axial force A GES can be generated and which supports the decoupling process of the axial coupling under load, covers an angular range of 3 degrees to 35 degrees.

[0065] Axial couplings are preferably used in two-wheelers that are exclusively pedal-driven or manually driven, or in two-wheelers that are additionally or exclusively motor-driven, especially in e-bikes or bicycles and cargo bikes that are assisted by an electric motor.

[0066] 5. Brief description of the drawings

[0067] In the following, exemplary embodiments of the invention and the derivation of the underlying dimensioning method are described with reference to the figures. These figures show: Fig. 1a: a prior art axial coupling with straight toothed

[0068] Plug-in gearing and orthogonal coupling gearing,

[0069] Fig. 1 b: an axial coupling according to the invention with helical toothed

[0070] Plug-in toothing and active surfaces of the coupling teeth tilted relative to an imaginary plane through the rotation axis,

[0071] Fig. 1c: an orthogonal view in the direction of rotation axis with a

[0072] cutout marking,

[0073] Fig. 1d: the section from Fig. 1c supplemented by the sketch of a blunt double wedge as a model for the movable coupling part, which is used to derive the design methods,

[0074] Fig. 1e: a model of the clutch assembly from Fig. 1 b, supplemented by a force diagram for a torque that would rotate the sun gear clockwise when viewed from the left,

[0075] Fig. 2: the schematic transmission section of a rear wheel hub from the

[0076] DE 10 2021 129423 A1 ,

[0077] Fig. 3a: the installation situation of a KHR coupling of a preferred

[0078] Embodiment of the invention,

[0079] Fig. 3b: a preferred embodiment of a control groove NHR for the

[0080] case of undercompensation,

[0081] Fig. 3c: another preferred embodiment of a control groove NHR for the case of compensation or overcompensation of the axial forces,

[0082] Fig. 3d: the installation situation of a KSR coupling, which is indirectly connected by means of a

[0083] Clutch KHR is automatically switched via inclined running and / or plug-in teeth,

[0084] Fig. 4a: a preferred embodiment of a switchable under load

[0085] Coupling K70, which is a block-type circulating coupling,

[0086] Fig. 4b: a preferred embodiment of a development of the control groove

[0087] N70 on the switching drum ST and the detent positions KR of the switching positions G1-G10,

[0088] Fig. 4c: a preferred embodiment of a control groove N70 for the case of undercompensation, Fig. 5a: a preferred embodiment of an installation situation of a

[0089] Coupling K40 in an isolated and more detailed representation with compensation or overcompensation of the axial forces,

[0090] Fig. 5b: a preferred embodiment of a control groove N40 for the case of undercompensation,

[0091] Fig. 5c: a preferred embodiment of a control groove N40 for the case of compensation or overcompensation,

[0092] Fig. 6: a preferred embodiment of an installation situation of a

[0093] Clutch K90 and

[0094] Fig. 7: a preferred embodiment of a section of a

[0095] Input gear EGG with helical gear teeth.

[0096] 6. Detailed description

[0097] Embodiments of the present invention are described below by way of example only. These examples represent the best ways of practicing the invention currently known to the applicant, although they are, of course, not the only ways in which this could be achieved. The description sets forth the functions of the example and the sequence of steps for designing and operating the example. However, the same or equivalent functions and sequences may be achieved by other examples.

[0098] Identical components have the same reference numerals.

[0099] Fig. 1a shows a prior art axial coupling with straight-toothed plug-in gearing and orthogonal coupling gearing.

[0100] Fig. 1b shows an axial coupling according to the invention with helical splines and the active surfaces of the coupling teeth tilted relative to an imaginary plane through the rotational axis. A first shaft 200 applies a torque M1 to the axial coupling 100. The first shaft 200 is connected either via a spline (not shown) or rigidly to a spur gear 120.

[0101] In the coupled state of the axial coupling 100, a second shaft 220 takes over a torque M2 which is equal in magnitude to M1.

[0102] The first shaft 200 and the second shaft 220 are mounted in a gear arrangement, in principle rotatable about a rotational axis 210, both in the case of a spur gear and in the case of a planetary gear. However, in the case of a planetary gear, the first shaft 200 or the second shaft 220 can also be axially and rotationally fixed.

[0103] The axial coupling 100 comprises a coupling ring 130, which, in the engaged state, engages a coupling toothing 114, which is connected in a rotationally fixed manner or alternatively in one piece with a spline 113. The spline 113 is axially displaceable and rotationally fixedly connected to the spur gear 120. The spline 113 comprises at least one tooth 112, which has a helix angle β S from 0 degrees to 30 degrees.

[0104] The clutch ring 130 comprises at least one asymmetrical tooth 131, which, when engaged, engages with a likewise asymmetrical toothing 114 of the movable clutch part 110. The asymmetrical teeth 131, 114 have a steep flank 132 and a flatter ramp 133 (see Fig. 1d). The ramp 133 has a helix angle β3 in the range of 45 degrees to 75 degrees with respect to an imaginary plane through the rotation axis 210 (see Fig. 1d). This serves to reliably overcompensate the frictional forces so that the clutch can operate as a freewheel. The flank 132 has a helix angle β with respect to an imaginary plane through the rotation axis 210. K in the range from 0 degrees to a maximum of 30 degrees (see Fig. 1d). The two helix angles β S and β K or the two helix angles β L and β K or the two helix angles β L and β Sare each in total 3 degrees to 35 degrees, in particular 5 degrees to 22 degrees and preferably 7 degrees to 20 degrees and provide the axial forces A required to support the separation of the axial coupling 100 under load S and A K (see Fig. 1e). The axial forces A S and A K support the opening of the axial coupling 100 in the opening direction x as soon as the coupling ring 130 is controlled in the opening direction x by an actuator 140 (see Fig. 1 e).

[0105] The sum of the two helix angles βS plus βK or βL plus βK or βL plus βS has a value in the range between 3 degrees and 35 degrees, in particular in the range between 5 degrees and 22 degrees, preferably in the range between 7 degrees and 20 degrees.

[0106] The axial force components A S and A K help the friction forces R S and R KThis means that undercompensation, compensation or overcompensation of the friction forces R S and R K The axial forces A S and A K are dependent on the difference between the amount of the input torque acting on the axial coupling 100 of the respective gear stage to be disengaged and the amount of the output torque of the respective gear stage to be disengaged.

[0107] The axial forces A S and A K support the opening of the axial coupling 100 in the opening direction x as soon as it is controlled by an actuator 140 (see Fig. 1e) in the opening direction x.

[0108] The preferred materials are steel and aluminum or steel and steel. However, other materials in combination with steel or in combination with each other, such as aluminum bronze, bronze, beryllium copper, brass, industrial ceramics, or solid or surface-coated components with a low-friction but tribologically resistant surface, for example made of carbon or carbon fiber, can also be considered in order to achieve the smallest possible friction angle ρ. S and ρ K to be able to realize designs in which a safe coupling and uncoupling between the movable coupling part 110 and the coupling ring 130 can be ensured even under high loads and strong external vibrations.

[0109] The axial coupling comprises a material pairing between the coupling ring 130 and the coupling toothing 114 and between the spline 113 and the guide of the spur gear 120 a material pairing of steel, aluminum, aluminum multi-component bronze, bronze, beryllium copper, brass, industrial ceramics or solid or surface-coated components carbon or carbon fibers in order to ensure safe coupling and decoupling between the lock and the coupling ring.

[0110] A force closing the clutch is exerted on the movable coupling part 110 via optional springs 111; 111' relative to a fixed housing part or relative to an axially rigid component (see Fig. 1e).

[0111] The movable coupling part 110 is guided axially movable via the spur gear 120. The spur gear 120 has a spline with a helix angle β S on.

[0112] Fig. 1c shows an orthogonal view in the direction of the rotation axis with a cutout marking.

[0113] Fig.1d shows the section from Fig. 1c supplemented by the sketch of a blunt double wedge as a model for the movable coupling part, which is used to derive the design methods.

[0114] On the movable coupling part 110 with spline, the helix angle β is shown for the example of a tooth 112 of the spline. S and using the example of a tooth of the coupling toothing 114, the helix angle β K which together form a derived axially movable wedge model, which is shown in more detail in Fig. 1e.

[0115] Fig. 1e shows a model of the clutch assembly from Fig. 1b, supplemented by a force diagram for a torque that tends to rotate the sun gear clockwise when viewed from the left. This shows the double wedge, which represents a problem analogous to an axial coupling with a spline and a clutch toothing. In this analogy, the wedge is the movable part of the clutch. It has two helical toothed active surfaces, for which the mathematical derivation is analogous (i = S for the spline and i = K for the clutch toothing). All forces are considered in terms of magnitude below. The normal forces N act on the active surfaces. i This results in frictional forces acting parallel to the tooth flanks:

[0116] R i = μ i * N i / 1 /

[0117] The tangential component of the normal force, Ni * cos(β i ), must be determined by the applied tangential force F iIn addition, the friction force also has a component R i * sin(β i ) in the direction of the applied tangential force, which it must also exert. Therefore, the following applies:

[0118] F i = N i * cos(β i ) + R i * sin(β i ) / 2 /

[0119] Using equations / 1 / and / 2 / the following relation can be derived:

[0120] N i = F i / (cos(β i ) + μ i * sin(β i )) / 3 /

[0121] The applied tangential force F i depends on the applied torque and the effective pitch radius of the gearing:

[0122] F i = M / r i / 4 /

[0123] In the axial direction, a component of the normal force, A i = N i * sin(β i ), as well as a component of the friction force, R i * cos(β i) in the opposite direction. The resulting axial force X i is therefore: X i = N i * sin(β i ) - R i * cos(β i ) / 5 / Using equations / 1 / to / 5 / , the axial force X i to X i = M / r i * [ sin(β i ) - μ i * cos(β i ) ] / [ cos(β i ) + μ i * sin(β i ) ] / 6 / For the special case A i = 0 we get: β1= arctan(μ i ) / 7 /

[0124] A helix angle that satisfies this relation is also called friction angle ρ i called.

[0125] The total axial force acting on the moving part of the coupling is the sum of the axial forces of the plug-in and coupling teeth, i.e.:

[0126] X GES = X S + X K / 8 /

[0127] According to the invention, in the case of complete compensation of the friction forces, the sum of the axial forces should be zero, i.e. X GES = 0. This means:

[0128] [ sin(β K ) - μ K * cos(β K ) ] / [ cos(β K ) + μ K * sin(β K ) ]

[0129] = - r K / r S * [ sin(β S ) - μ S * cos(β S ) ] / [ cos(β S ) + μ S * sin(β S ) ] / 9 / β K is therefore dependent on β S , where a solution can be found by: β K = 2*arctan( [ a*b + ((a 2 + 1 )*(b 2 + 1 )) 1 / 2 - 1 ] / [ a + b ] ) / 10 / with: a = μ K b = - r K / r S * [ sin(β S ) - μ S *cos(β S ) ] / [ cos(β S ) + μ S *sin(β S ) ]

[0130] Empirically, a sufficiently good approximation could be determined for common helix angles, given by: β K = ρ K - (r K / r S ) * (β S - ρ S ) / 11 /

[0131] For the further embodiments according to the invention comprising an inclined position of a plug-in toothing and a running toothing or of a running toothing and a coupling toothing, the derivation is analogous.

[0132] In the following, examples of specific embodiments of the invention are described, as they are used within an overall device such as a clutch assembly or a transmission system. Fig. 2 shows a rear wheel hub with a multi-speed transmission formed by planetary gears that engage in spiral grooves N24 to N74, NSR, NHR in a shift drum by means of axially displaceable sliding rings with shift fingers. The shift fingers are also guided in radial slots oriented parallel to the transmission axis of a hollow main axis HA surrounding them, and various gears can be shifted by means of the sliding rings, which couple and spring-load radially extending clutch rings equipped with axial teeth according to the invention are released according to a respective gear stage. The multi-speed transmission arranged on the main axis consists of a five-speed transmission as the input transmission EGG, starting from a drive sleeve AH.which consists of a two-stage and a subsequent single-stage planetary gear, as well as a fast-transforming, subsequent single-stage planetary gear as a secondary transmission NSG, which is arranged on a central sleeve MH and, when coupled, connects the first five gears to a hub sleeve NH or, with a fixed ring gear HR4, transfers these gears to the hub sleeve NH in a double-coupled manner, whereby these 10 gear stages thus formed are switched with 7 clutches K20, K30, ..., KHR in such a way that only a few of the planetary gear sets are engaged in a rolling manner under load.

[0133] At least two of the planetary gears PR1, PR2 are identical to each other and have interconnected annular ring gears HR2-3, and the planet carriers PT2, PT3 of these two planetary gears are axially and radially supported against each other by an intermediate bearing ZL.

[0134] The sun gear SR5 of the secondary transmission NSG is mounted on the center sleeve MH and is connected to the hub sleeve NH for axial displacement and rotational stability. It can be coupled to the output-side planetary carrier PT3 of the input transmission EGG using a sun gear clutch KSR, according to the invention. The ring gear HR4 of the secondary transmission NSG can be coupled to the main axle HA using a ring gear clutch KHR, according to the invention. Fig. 3a shows the installation situation of a KHR clutch of a preferred embodiment of the invention.

[0135] Fig. 3b shows a preferred embodiment of a development of the control groove NHR on the switching drum ST and the detent positions KR of the switching positions G1-G10.

[0136] Fig. 3c shows a preferred embodiment of a control groove NHR for the case of compensation or overcompensation of the axial forces.

[0137] Fig. 3d shows a further simplified embodiment of the rear-mounted transmission in which the sun gear clutch KSR is not engaged from the shift drum ST, but rather the thrust forces of a helical gearing of the rear-mounted transmission and / or the helical gearing of the torque transmitter's plug-in gearing in the hub sleeve NH are utilized such that the sun gear clutch KSR is always automatically opened or closed against a spring force, opposite to the actively engaged ring gear clutch KHR. Accordingly, in direct gear, the sun gear clutch KSR is closed and the ring gear clutch KHR is open. The rear-mounted transmission therefore rotates as a single unit.In the preferred variant, the running gear helix and / or the spline helix are oriented such that the sun gear SR5 moves away from the stationary part of the KSR clutch under load due to the axial force, and the ring gear HR5 experiences an opposing force, which is supported by the axial support bearing ASL against the NH hub sleeve. The torque transmitter D5 is connected to the NH hub sleeve in a rotationally fixed but axially movable manner, and is also connected to the sun gear SR5 in a rotationally fixed and axially fixed manner. When stationary or during operation in direct gear of the NSG rear-mounted transmission, the KSR sun gear clutch is closed by the return spring F5, which axially supports the torque transmitter D5 against the NH hub sleeve. This movement is limited by a stop in the NH hub sleeve.The axial forces of the HR5 ring gear are absorbed directly by an axially loadable roller bearing between the NH hub sleeve and the HR5 ring gear, and / or via an axial support element ASE on the main axle HA. The shifting process is supported by the fact that when the KHR ring gear clutch is engaged, the SR5 sun gear overtakes the PT5 planet carrier and thus switches to freewheel mode. The orientation of the clutch teeth also ensures that the KSR clutch is released. The drive of the NH hub sleeve is shifted from the MH center sleeve to the SR5 sun gear without any interruption in the load.

[0138] Fig. 4a shows a preferred embodiment of a clutch K70 according to the invention that can be switched under load. It is a block-type recirculating clutch that connects or disconnects two rotating parts by switching. It is, for example, the clutch K70 of a 10-speed hub gear from patent application DE 10 2021 129 423.1. This clutch is designed according to the invention here. The reference numerals used there are also used and supplemented here. The clutch K70 designed according to the invention in this system example connects the sun gear SR3 to the center sleeve MH, which also serves as a partial web for the output-side part of the input gear. The movable part of the clutch is guided axially movable and rotationally secured inside the center sleeve by means of a spline. The axially fixed part of the clutch is housed in one piece on the output side of the axially closely guided sun gear SR3.The movable part of the clutch has an annular collar on the inside which serves as a contact surface for the sliding ring which is guided in a rotationally secured manner so that it can move axially and is guided in an axially movable manner on the main axis. The sliding ring is designed in such a way that the collar is freely rotating inside the movable clutch part and is closely guided by a rotationally secured, spring-loaded stop ring and an axially stepped stop surface of the sliding ring. The axial length of the step is only slightly greater than the wall thickness of the inner annular collar of the movable clutch part. This prevents a clamping effect. The spring F71, which acts in the opening direction, secures the open position of the clutch. The spring F70, which acts in the closing direction, supports the closing process of the clutch because a track change takes place in the control groove N70 on the shift drum ST.If the static friction is only partially compensated, the sliding ring and thrust washer are installed rotated 180 degrees, and the spring F71 is omitted. This is because, supported by the axial forces, the clutch K70 is now opened against the spring force of the F70. This also applies to the other examples.

[0139] With the K70, a helical spline and a helical coupling spline can be combined. The coupling spline is preferably helical.

[0140] The axial forces A S and A K (see Fig. 1e) support the opening of the axial coupling 100 in the opening direction x (also shown in Fig. 1e) as soon as it is controlled by an actuator 140.

[0141] Fig. 4b shows a preferred embodiment of a development of the control groove N70 on the switching drum ST and the detent positions KR of the switching positions G1-G10 in the case of undercompensation.

[0142] Fig. 4c shows a preferred embodiment of a control groove N70 for the case of compensation or overcompensation. The shift finger SF is tightly guided by both flanks of the control groove N70.

[0143] Fig. 5a shows a preferred embodiment of an installation situation of a clutch K40 according to the invention in an isolated and more detailed representation with compensation or overcompensation of the axial forces. In this embodiment, a helical spline and / or a helical clutch spline is used. It is a planetary stage whose input is a planetary carrier and the output is via a ring gear. To change gears, a sun gear SR2 is engaged or released in a rotationally fixed manner. A follower lock is required for this. When releasing the sun gear SR2, the clutch K40 must be opened under load. Within the sun gear SR2, the movable clutch part is guided concentrically to it by means of a spline SST, which is rotationally fixed and axially displaceable. To support the shifting process, the spline SST is provided with a helix angle β Sand / or the effective surfaces of the clutch teeth relative to an imaginary plane through the rotation axis of the gearbox by the angle β K twisted. The non-rotatable and axially fixed part of the K40 coupling is connected to the main axis HA via a spline SST. The axial locking is preferably achieved, as shown in Fig. 5a, by means of screws that engage in holes adapted to the screw. Fig. 5c shows the control of the axial coupling for the case of compensation and overcompensation of the friction forces. The control of the K40 coupling corresponds to that of the K70 coupling. In the case of undercompensation, the description there also applies here.

[0144] The axial forces A S and A K (see Fig. 1e) support the opening of the axial coupling 100 in the opening direction x (also shown in Fig. 1e) as soon as it is controlled by an actuator 140.

[0145] Fig. 5b shows a preferred embodiment of a groove N40 for the case of undercompensation.

[0146] Fig. 5c shows a preferred embodiment of a groove N40 for the case of compensation or overcompensation.

[0147] Fig. 6 shows a preferred embodiment of an installation situation for a clutch K90. A preferred embodiment of the present invention for such a transmission is shown schematically in Fig. 6. The output-side clutch K90 of the rear-mounted transmission NSG shown there is controlled via two cascaded shift fingers, one of which is guided in a groove N94 in the shift drum ST (play arm) and the second of which passes through the rotating output sleeve H and controls a clutch ring 130. Two springs F90 and F91 center the stacked shift fingers. When the clutch K90 is engaged, the output torque is transferred from the output-side side wall of the ring gear HR to the output sleeve, which can rotate freely because a retaining ring SR defines its axial play on the output sleeve H.

[0148] According to another preferred embodiment, an axial clutch 100 is positioned closer to a main axis. The axial clutch 100 is located in a rotating sleeve H. A helical gear used to facilitate shifting is arranged between a planetary carrier PT and a clutch half on the output side. Shift fingers SF and centering springs also engage this half. Stacking of shift fingers SF is thus eliminated, which significantly simplifies design and manufacturing.

[0149] In the clutches in the transmission according to DE 10 2018 007 A1, the shifting forces of some K50 and K90 clutches are transmitted to the shift rings via several stages in the output-side transmission block. Shifting elements pass through rotating sleeves. Therefore, if the gear teeth are tilted, axial forces on the shift elements occur that are difficult to control or are required for shifting. Therefore, only slight inclinations of a few degrees are provided.

[0150] The axial forces A S and A K(see Fig. 1e, respectively) support the opening of the axial clutch 100 in the opening direction x (also shown in Fig. 1e), as soon as it is controlled by an actuator 140. Fig. 7 shows a section of the input transmission EGG with at least one axial clutch 100 according to the invention. The two symmetrical planetary gears with the planet gears PR2 and PR3 are provided with symbolically represented helical gears. The resulting axial forces are directed as indicated by the arrows on the sun gears SR2, SR3 and on the ring gears HR2 and HR3, namely in the opening directions of the clutches K40, K50 according to the invention. The reaction forces acting via the planet gears PR2, PR3 on the interconnected ring gears HR2 and HR3 compensate each other and keep them centered on one another as a so-called flying ring pair. The opening forces A S and A KClosing springs F40 and F50 are directed in opposite directions. This requires relatively moderate shifting forces on the shift fingers, which each engage with a loose fit in control grooves N40 and N50, to accomplish the switching of the axial clutches even under high transmission loads.

[0151] The release springs F41, F51 are dimensioned so that the sliding rings secure open clutch positions, even if no auxiliary force from the helical gearing is available due to lack of or low gear load.

[0152] It is advantageous that the sun gears and the associated clutch ring are manufactured as a single piece. In the rear-wheel transmission described, a clutch travel of only approximately 1.5 mm between the engaged and disengaged positions is sufficient.

[0153] In Fig. 7, the axial forces of the running gear and those of the clutch gear are utilized. Only clutch K40 is opened under load. Clutch K50 enters freewheel mode when clutch K70, the block-type recirculating clutch, is closed. Sun gear SR3 is then simply driven in the direction of rotation. Therefore, the helical gearing of clutch K50 could be omitted. The control then remains as shown in Fig. 2. This provides a design that enables effective solutions in tight space conditions. In this case, the axial forces of clutch K50 are undercompensated.

[0154] The K40 clutch works as a follow-on lock and the K50 clutch works as a backstop.

[0155] In a preferred embodiment, the helical gearing is located on the running gear teeth and also in the clutch gear teeth. This is due to the geometric and tribological situation, since only the K40 clutch needs to be engaged under load. Assuming a rolling friction coefficient of the running gear teeth of μ L = 0.015, which corresponds to a friction angle ρ L of 0.86 degrees, based on a pitch circle radius r L of 25 mm, a mean pitch circle radius r K the clutch teeth of 17.2 mm with a friction coefficient μ K of 0.12, which corresponds to a friction angle ρ K of 6.84 degrees, the following design variants result when compensating the friction forces according to the inventive method: β K = 0 => β L = 10.9 degrees, β K = ρ K = 6.84 degrees, β L = ρ L = 0.86 degrees, β K = 3 degrees => β L = 6.44 degrees and βK = 5 degrees => β L = 3.54 degrees.

[0156] It can be seen that the sum of the friction angles ρ K and ρ L only corresponds to the actually required angle sum for individually compensated gears.

[0157] The axial force A GESsupports the opening of the axial coupling 100 in the opening direction x as soon as it is controlled by an actuator 140 (see Fig. 1e) in the opening direction x.Figures 2 to 7 show a clutch assembly or sections of a clutch assembly comprising a planetary gear with at least one axial clutch 100 with an axially displaceable clutch ring 130 for actuating an axially toothed axial clutch 100, which on the one hand is subjected to a closing force by a closing spring 111, 111' and also with at least one actuator 140, comprising at least one shift finger engaging in a partially coiled guide groove of a shift drum, so that in predetermined rotational positions of the same, the groove side wall thereof exerts a shift force against the closing force on the shift fingers and thereby opens the axial clutch 100 and thus produces a different shift state of the respective gear stage, characterized in that the toothing of the planetary gear is designed as a helical toothing such that torque-dependent axial forces A occurring under a load. GESin the opening direction of the axial clutch 100, comprising the clutches K40, K50, and largely compensate for the static friction occurring between the contacting clutch teeth, which counteracts a separation of the axial clutch 100, wherein the resulting reaction forces occurring on the clutch elements are absorbed by support elements on adjacent transmission components.

[0158] The list of reference symbols refers to the attached figures.

[0159] 100 axial coupling

[0160] 110 movable coupling part

[0161] 111 ; 111' spring

[0162] 112 at least one tooth of the spline

[0163] 113 spline

[0164] 114 Clutch teeth

[0165] 120 spur gear

[0166] 130 coupling ring

[0167] 131 asymmetric teeth

[0168] 132 steep flank of the clutch teeth

[0169] 133 flat ramp of the clutch teeth

[0170] 140 Actuator

[0171] 200 first wave

[0172] 220 second wave

[0173] 210 Rotation axis

[0174] 300 sliding ring with collar

[0175] 600 running gear

[0176] WP1 first active surface pairing

[0177] WP2 second active surface pairing

[0178] MP1 first material pairing

[0179] MP2 second material pairing ρ S Friction angle spline ρ K Friction angle of clutch teeth ρ L Friction angle of running gear β S Helix angle of 113 β K Helix angle of 114 β L Helix angle of 600 β1first helix angle β2second helix angle β3helix angle of 133 μ S Friction value of spline μ KFriction value of clutch teeth μ L Friction value of running gear r S Pitch circle radius spline r K mean pitch circle radius of coupling teeth r L Pitch circle radius running gear

[0180] N S , N K Normal force plug-in / coupling toothing

[0181] R S , R K Friction force plug-in / coupling teeth

[0182] F S , F K Tangential force plug-in / coupling toothing

[0183] A S , A K axial component of N S , N K A GES Sum of the angled

[0184] gears

[0185] Generated axial forces

[0186] X S , X K Effective axial force at

[0187] Plug-in / coupling toothing

[0188] X GES Sum of effective axial forces x, opening direction

[0189] M1 , M2 torque

[0190] M Amount of torque M1 / M2

[0191] AH drive sleeve

[0192] HA main axis

[0193] MH center sleeve

[0194] NH hub shell

[0195] ST shift drum

[0196] N20, N30, N40, N50, N60, N70,

[0197] N80, N90, NSR, NHR control groove(s)

[0198] L bearings from MH

[0199] PT1-2, PT3 / PT3-4, PT5 planet carrier

[0200] ZL intermediate storage PT1-2 to PT3-4

[0201] SR1 - SR5 sun gears

[0202] PR1 - PR5 planetary gears

[0203] HR2-3, HR5 ring gears F5 return spring

[0204] F40, F50 closing spring

[0205] F41, F51 release spring

[0206] F70, F71 spring

[0207] ASL axial support bearing

[0208] ASE axial support element

[0209] D5 torque transformer

[0210] KR locking positions

[0211] AS thrust washer

[0212] IGH Integrated Gear Hub

[0213] HR ring gear

[0214] PR planetary gear

[0215] SR sun gear

[0216] EGG input gear

[0217] NSG rear-mounted gearbox

[0218] G1... G10 switching positions

[0219] K20, K30, K40 clutches first sub-gearbox

[0220] K50, (K60), K70, K80 clutches second sub-gearbox

[0221] KSR sun gear clutch in the NSG

[0222] KHR ring gear coupling in the NSG

[0223] SST helical spline

[0224] SF shift finger

[0225] K backdrop

[0226] H sleeve

[0227] Symbol list

Claims

Claim 1. An axial coupling (100) comprising a first shaft (200) and a second shaft (220) which can be coupled and uncoupled controlled by an actuator (140), comprising a first working surface pair (WP1) with a first helix angle (β1) and a second working surface pair (WP2) with a second helix angle (β2), wherein the first helix angle (β1) comprises an angular range of 0 degrees to 30 degrees with respect to an imaginary plane through an axis of rotation (210) and the second helix angle (β2) comprises an angular range of 0 degrees to 30 degrees with respect to an imaginary plane through an axis of rotation (210), and the sum of the helix angles (β1) plus (β2) over which a torque-dependent axial force (A) is applied. GES) is generable and which supports the uncoupling process of the axial coupling (100) under load, encompasses an angular range of 3 degrees to 35 degrees.

2. Axial coupling (100) according to claim 1, wherein the first effective surface pairing (WP1) comprises a splined connection (113) with at least one tooth (112) and the first helix angle (β1) has a helix angle (β S ) comprises and the second functional surface pairing (WP2) comprises a coupling toothing (114) and the second helix angle (β2) comprises a helix angle (β K ) comprises.

3. Axial coupling (100) according to claim 2, wherein a coupling ring (130) of the coupling toothing (114) has an asymmetric toothing (131 ) with at least one tooth (112), comprising a flank (132) the angle (β ). K ) has.

4. Axial coupling (100) according to claim 3, wherein the coupling ring (130) is designed to be controllable in an opening direction x by an actuator (140), wherein the axial force (A GES) the opening of the axial coupling (100) in the opening direction x, supported.

5. Axial coupling (100) according to claim 1 wherein the first The effective surface pairing (WP1) comprises a splined connection (113) with at least one tooth (112) and the first helix angle (β1) has a helix angle (β S ) includes and the second functional surface pairing (WP2) includes a running gear (600) and the second skew angle (β2) a skew angle (β L ) comprises 6. axial coupling (100) according to claim 1 wherein the first The effective surface pairing (WP1) comprises a coupling toothing (114) with at least one tooth (112), and the first helix angle (ßi) is a helix angle (β K ) includes and the second contact surface pairing (WP2) includes a running gear (600) and the second helix angle (β2) includes a helix angle (β L) comprises 7. Axial coupling (100) according to any one of the preceding claims, wherein the first material pairing (MP1) of the first contact surface pairing (WP1) and the second material pairing (MP2) of the second contact surface pairing (WP2) comprise one of the material pairings of steel, bronze, beryllium copper, aluminum, aluminum multi-component bronze, ceramic, carbon or carbon fibers individually or in combination to ensure safe uncoupling. 8.Clutch assembly comprising a planetary gear with at least one axial clutch (100) with an axially displaceable clutch ring (130) for actuating an axially toothed axial clutch (100), which on the one hand is acted upon by a closing spring (111 , 111') with a closing force and in addition with at least one actuator (140), comprising at least one switching finger engaging in a partially helical guide groove of a switching drum, so that in each predetermined rotational position of the same its groove side wall exerts a switching force against the closing force on the switching fingers and. thereby opening the axial clutch (100) and thus establishing a different switching state of the respective gear stage, characterized in that the gearing of the planetary gear is designed as a helical gear such that torque-dependent axial forces (A) occurring under load GES) in the opening direction of the axial coupling (100), comprising the couplings (K40, K50), and largely compensate for the static friction occurring between the contacting coupling teeth, which acts in the opposite direction to a separation of the axial coupling (100), whereby the resulting reaction forces occurring at the coupling elements are absorbed by support elements on adjacent transmission components.

9. Coupling assembly comprising a planetary gear unit according to claim 8, characterized in that sun gears (SR2, SR3) of two mirror-image adjacent helical planetary gear units, which are connected to a main shaft (HA) by couplings (K40, K50) orfrom this, whereby the respective reaction forces occurring in the sun gears, which are transmitted to the associated planet gears (PR2, PR3) and from these to the associated ring gears (HR2-3), are mutually absorbed by their interaction.

10. Coupling assembly comprising a planetary gear set according to claim 8, characterized in that the coupling rings are each connected to the coupling (K40, K50).

11. Coupling assembly comprising a planetary gear set according to any one of claims 8 to 10, characterized in that the helical gearing is a running gear set whose helix angle is between 5 degrees and 12 degrees, in particular between 8 degrees and 10 degrees.

12. Clutch assembly comprising a planetary gear set according to claim 8, characterized in that the helical gearing is a splined connection and the ring gear is supported by a support bearing on the hub sleeve and above it on the gear housing.

13. Clutch assembly comprising a planetary gear set according to claim 8, characterized in that the shift finger, which is guided in the groove (N60 / N80) in the control drum (ST) with minimal tolerance, is held axially centered on the main shaft (HA) on both sides between two springs, each of which is axially fixed at its end end by support discs on the main shaft, axially drives a sun gear with helical teeth towards the planet carrier (PT), which carries a clutch ring, the counterpart of which is axially fixed on the rotatable sleeve (H) and forms the clutch (K60, K80) to be shifted.

14. Method for dimensioning the axial clutch according to one of the preceding claims in the following steps: • Determination of a coefficient of friction (μ1) between the pair of contact surfaces (WP1) involved in a first material pairing (MP1) and a coefficient of friction (μ2) between the pair of contact surfaces (WP2) involved in a second material pairing (MP2) using standard methods on the object and / or taken from a table of material properties. • Determination of an effective partial circle radius (r1) of the first functional surface pairing (WP1 ) and an effective partial circle radius (r2) of the second functional surface pairing (WP2) from the design data. • Determination or specification of a first helix angle (β1) of the first functional surface pairing (WP1) and a second helix angle (β2) of the second functional surface pairing (WP2) from such design data, which are load-independent, namely the coefficients of friction of the first material pairing (MP1) and the second material pairing (MP2) (μ1) and (μ2), the pitch circle radii of the first contact surface pairing (WP1) and the second contact surface pairing (WP2) (r1) and (r2), where, for example, in the case of complete compensation of the frictional forces, the angle (β2) can be determined as a function of (β1): β2 = 2 * arctan( [ a * b + ((a 2 + 1 )*(b 2 + 1 )) 1 / 2 - 1 ] / [ a + b ] ) with: a = μ2b = - r2 / r1* [ sin(β1) - μ1*cos(β1) ] / [ cos(β1) + μ1*sin(β1) ].

15. Method for dimensioning the axial coupling according to claim 14, comprising an approximation for angles up to 30° with: β2= ρ2- (r2 / r1) * (β1- ρ1) and with: ρ i = arctan(μ i ).

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