Improved shifting of axial couplings with flat teeth under load

Abstract

Axial coupling (100) with a first shaft (200) and with a second shaft (220), which can be coupled and uncoupled controlled by an actuator (140), comprising a first working surface pairing (WP1) with a first helix angle (β1) and a second working 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 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 generated and which supports the uncoupling process of the axial coupling (100) under load, encompasses an angular range of 3 degrees to 35 degrees, wherein the first effective surface pairing (WP1) and the second effective surface pairing (WP2) are each formed by a toothing, the toothings being a coupling toothing and a running toothing of spur gears of a planetary gear or a spur gear, or a plug toothing and a running toothing of spur gears of a planetary gear or a spur gear, and the torque-dependent axial force (A) is generated by the inclination of the toothings under drive load. GES ) can be generated, which acts in the opening direction (x) of the axial coupling (100) and supports its uncoupling process under load.

Description

1. Technical field

[0001] The present invention relates to an arrangement of planetary gears or spur gears that can be switched into different gears, 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 rotating the clutch ring for the purpose of gear selection, the clutch ring engages or disengages the respective clutch by means of a switching force occurring on the side wall of the groove, which is directed against the pressure force. 2. State of the art

[0002] Such transmissions with axial clutches for setting different gear ratios are 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 129 423 A1. All gears in the aforementioned transmissions can be shifted even under operating load. However, considerable shifting forces are required during individual shifts to disengage a clutch against the load-induced retaining forces that occur between the clutch elements, so that a reduction in the drive torque of the crankshaft and / or the auxiliary motor is advisable during the corresponding shifting operations. DE 10 2021 129 423 A1 describes a downstream transmission which has helical gear teeth used for the indirect shifting of the sun gear clutch KSR.The shifting process is triggered by the change in state of the KHR clutch. The sun gear clutch does not have an actuator controlled via a control groove in a shift drum. Furthermore, after the KHR ring gear clutch engages, the KSR clutch is in freewheel mode because the sun gear SR rotates faster than the ring gear.

[0003] DE 10 2014 101 726 A1 describes a power-shiftable multi-speed transmission for use in passenger vehicles, commercial vehicles, bicycles, etc.

[0004] From DE 11 2017 007 405 T5 a coupling mechanism for a bicycle or a bicycle with an additional electric drive is known.

[0005] DE 20 2012 009 449 U1 describes a movement device for a movable furniture part.

[0006] A switching device and a vehicle powertrain as well as a method for operating a switching device and a vehicle powertrain is known from DE 10 2013 213 156 A1. 3. Task definition |

[0007] The object of the present invention is to improve the switching capability of axial couplings of planetary gears or spur gears in such a way that they can be reliably coupled and uncoupled with low switching forces by an actuator, even under drive load. 4. Summary of the invention

[0008] The problem is solved according to the features of independent claim 1 and independent claim 5. Further advantageous embodiments of the solution are found in the respective dependent claims.

[0009] The solution lies in the fact that the gearing, in particular a) a coupling toothing and a running toothing of the spur gears or b) a splined connection and a running connection of the spur gears a planetary gear or a spur gear is / are designed with regard to the skew of the gear teeth in such a way that the torque-dependent axial forces occurring under drive load act in the opening direction of an axial coupling, and are largely compensated by the static friction acting against the disengagement of the axial coupling between the contacting coupling teeth and the contacting plug-in teeth or between the contacting coupling teeth and the contacting impeller teeth, or between the contacting plug-in teeth and the contacting impeller teeth, whereby the resulting reaction forces occurring on the coupling elements are absorbed in support elements on adjacent gear components.

[0010] The spur gear or planetary gear is arranged in a housing that is fixedly connected, or by means of a torque arm connected to an axle in a rotationally and axially fixed manner, to the frame of a vehicle driven via at least one wheel and movable over a roadway. The frame provides 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). For the purposes of this discussion, a two-wheeled vehicle driven via at least one of its wheels is preferably assumed.

[0011] The gearbox is positioned between two shafts. The drive can, for example, include a shaft driven by muscle power via a pedal and / or a shaft driven by a motor.

[0012] A shaft within the meaning of this invention is a component which can absorb or transmit torque.

[0013] Within the gearbox, at least one axial coupling is arranged, which couples an incoming shaft with an outgoing shaft in such a way that coupling or uncoupling between the incoming shaft and the outgoing shaft is controlled via an actuator.

[0014] The input and output shafts are rotatably mounted in the gearbox housing of the spur gear unit and, typically, in the gearbox housing of the planetary gear unit. In the case of the planetary gear unit, one of the shafts may be fixed to the gearbox housing or the main shaft.

[0015] To improve the switching behavior, torque-dependent axial forces are used. a) a coupling toothing with a helix angle β Ktilted surface normals of the effective surfaces towards the axis of rotation, so that the angle between the effective surface and the back flank of the coupling teeth is increased or b) a helical gear with a helix angle β L or c) a helical splined connection with a helix angle β S generated.

[0016] If the respective inclination angle β K or β L or β S Since the axial force is equal to the arctangent of the coefficient of friction µ of the opposing contact surface pairs in the respective material pairing, the axial force corresponds in magnitude to the axial component of the frictional force. This angle is subsequently referred to as the friction angle p.

[0017] The friction angle ρ of the contact surface pairs depends on the material pairing of the opposing coupling materials in the axial coupling and the lubrication between the contact surface pairs. In principle, the temperature at the axial coupling also has an influence, but this can be neglected within the temperature range of typical gearbox applications. Another factor is the time the contact surface pairs are under load. A contact surface pair is defined as the surfaces that transmit the forces generated by the applied torque from the driving coupling part to the driven coupling part when the coupling is engaged.

[0018] 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 splined connection, whereby the axial mobility of the component carrying the coupling teeth is achieved by the axial displacement of the component with the coupling teeth relative to an axially tightly guided, driving component, as inherent in the splined connection, then the axial coupling can be opened and closed by means of the provided actuator. This enables potential rotation of the coupled shaft by the flat-toothed axial coupling.The axial force required to open the axial coupling and overcome the friction between the contact surfaces of the splined coupling and the contact surfaces of the flat-toothed coupling depends, in addition to the torque to be transmitted, on the material pairing of the contact surface pair of the splined coupling, their pitch circle diameter, and on the material pairing of the contact surface pair and the mean pitch circle diameter of the coupling. The mean pitch circle radius of the coupling is understood to be the mean radius of the circular ring that just completely encloses the radially oriented coupling teeth.

[0019] If the axially movable part of an axial coupling is provided with the degree of freedom necessary for engagement with the axially fixed coupling part by means of a toothed gear, whereby the axial mobility of the gear carrying the coupling teeth is achieved by the axial displacement of the gear with the coupling teeth relative to an axially tightly guided, driving gear, as inherent in the toothed gear, then the axial coupling can be opened and closed by means of the provided actuator. This enables potential rotation of the coupled shaft by the flat-toothed axial coupling.The axial force required to open the axial coupling in order to overcome the friction between the effective surfaces of the running gear and the effective surfaces of the flat-toothed coupling gear depends, in addition to the torque to be transmitted, on the material pairing of the effective surface pair of the running gear, their pitch circle diameter, and on the material pairing of the effective surface pair and the mean pitch circle diameter of the coupling gear.

[0020] If the axially movable part of an axial coupling is provided with the degree of freedom necessary for engagement with the axially fixed coupling part by means of a running tooth and a splined connection, whereby the axial mobility of the gear carrying the coupling teeth is achieved by the axial displacement of the gear with the coupling tooth and the splined connection relative to an axially tightly guided, driving gear, then the axial coupling, which opens indirectly via an actuator when entering freewheel mode and closes by means of a return spring when the drive torque of the driving gear is removed, allows for potential rotation of the disengaged shaft relative to the previously coupled shaft.

[0021] The axial force required to open the axial coupling and overcome the friction between the contact surfaces of the running gear and the contact surfaces of the splined shaft depends, in addition to the torque to be transmitted, on the material pairing of the contact surface pair of the running gear, their pitch circle diameter, and on the material pairing of the contact surface pair and the pitch circle diameter of the splined shaft. Since the axial coupling switches to freewheel mode when opening, no frictional forces occur between the contact surfaces of the coupling teeth. With this type of coupling, the drive shaft is changed under load.

[0022] The rule is: the smaller the radius, the greater the frictional force for a given torque.

[0023] In the variants according to the invention, there are two pairs of contact surfaces on which a first and a second frictional force is 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 effective pitch circle radii are generally different, so the different radii are also taken into account according to the invention.

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

[0026] The present disclosure also relates to a method for dimensioning the axial couplings according to the invention, taking into account the geometric and tribological conditions, comprising the following steps: • Determination of the coefficient of friction µ1 between the pair of contact surfaces WP1 involved in a first material pairing MP1 and the 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 from a table of material properties. • Determination of the effective partial circle radius r1 of the first functional surface pairing WP1 and the 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 contact surface pairing WP1 and a second helix angle β2 of the second contact surface pairing WP2 from such design data that are independent of the load, 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, whereby, 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+((a2+1)*(b2+1))1 / 2−1] / [a+b]) with: a=µ2 b=-r2 / r1*[sin(β1)-µ1*cos(β1)] / [cos(β1)+µ1*sin(β1)]

[0027] A sufficiently good approximation for angles up to 30° is achieved by: β2=ρ2−(r2 / r1)*(β1−ρ1) with: ρi=arctan(μi), i=1,2 i = 1,2

[0028] One simple solution is to compensate for the frictional forces individually, i.e.: β2=ρ2 and β1=ρ1.

[0029] In a first preferred and alternative embodiment of the axial coupling, it comprises, to support the uncoupling process of the axial coupling, a plug-in toothing in a first working surface pairing WP1 with material pairing MP1 and a coupling toothing in a second working surface pairing WP2 with material pairing MP2.

[0030] In this embodiment, the helix angle β1 is replaced by a helix angle β S The splined connection is formed. The helix angle β2 is determined by a helix angle β K formed by the coupling teeth.

[0031] The pitch circle radius r1 is formed by a pitch circle radius rs of the splined connection. The pitch circle radius r2 is formed by a mean pitch circle radius r. K the clutch teeth.

[0032] The coefficient of friction µ1 results from the first material pairing MP1 of the splined connection with a coefficient of friction µ S The coefficient of friction µ2 results from the second material pairing MP2 of the coupling teeth with a coefficient of friction µ. K .

[0033] In another alternative embodiment of the axial coupling, it comprises, to support the uncoupling process of the axial coupling, a splined connection in the first working surface pairing WP1 with material pairing MP1 and a running tooth in the second working surface pairing WP2 with material pairing MP2.

[0034] In this embodiment, the helix angle β1 is replaced by a helix angle β S The splined connection is formed. The helix angle β2 is determined by a helix angle β L formed by the running gear teeth.

[0035] The partial circle radius r1 is formed by a partial circle radius r Sof the splined connection. The pitch circle radius r2 is formed by a pitch circle radius r L the gear teeth.

[0036] The coefficient of friction µ1 results from the first material pairing MP1 of the splined connection with a coefficient of friction µ S The coefficient of friction µ2 results from the second material pairing MP2 of the running gear with a coefficient of friction µ. L .

[0037] In another alternative embodiment of the axial coupling, it comprises, to support the uncoupling process of the axial coupling, a coupling toothing in the first working surface pairing WP1 with material pairing MP1 and a running toothing in the second working surface pairing WP2 with material pairing MP2.

[0038] In this embodiment, the helix angle β1 is replaced by a helix angle β K the coupling teeth. The helix angle β2 is formed by a helix angle βL formed by the running gear teeth.

[0039] The partial circle radius r1 is formed by a mean partial circle radius r K of the coupling teeth. The pitch circle radius r2 is formed by a pitch circle radius r L the gear teeth.

[0040] The coefficient of friction µ1 results from the first material pairing MP1 of the coupling teeth with a coefficient of friction µ K The coefficient of friction µ2 results from the second material pairing MP2 of the running gear with a coefficient of friction µL.

[0041] The schematic drawings 1b - 1e show the relationships described below.

[0042] The axial coupling 100 allows a first shaft 200 to be coupled and uncoupled with a second shaft 220, controlled by an actuator 140, comprising a first working surface pairing WP1 with a first helix angle β1 and a second working surface pairing WP2 with a second helix angle β2, wherein the first helix angle β1 covers 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 covers 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, via which a torque-dependent axial force AGES can be generated and which supports the uncoupling process of the axial coupling 100 under load, covers an angular range of 3 degrees to 35 degrees.

[0043] In a first embodiment, the axial coupling 100 comprises a splined connection 113 with at least one tooth 112 in the first contact surface pairing WP1, and the first helix angle β1 comprises a helix angle β S and in the second contact surface pairing WP2 a coupling toothing 114 and the second helix angle β2 comprises a helix angle β K .

[0044] 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 forms the angle β K exhibits.

[0045] The coupling ring 130 is designed to be controllable in an opening direction x by an actuator 140, whereby the axial force A GES The opening of the axial coupling 100 in the opening direction x is supported.

[0046] In a further embodiment, the axial coupling 100 comprises in the first contact surface pairing WP1 a splined connection 113 with at least one tooth 112 and the first helix angle β1 comprises a helix angle β S and in the second contact surface pairing WP2 a running tooth 600 and the second helix angle β2 comprises a helix angle β L .

[0047] In a further embodiment, the axial coupling 100 comprises in the first working 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 contact surface pairing WP2 comprises a running gear 600 and the second helix angle β2 comprises a helix angle β L .

[0048] 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, multi-component aluminum 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., to ensure safe coupling and uncoupling.

[0049] The following examples describe preferred embodiments in detail. It is taken into account that the coefficients of friction depend on the material pairing, heat treatment (such as hardening), lubrication, contact duration, etc. This means that a significant variation in the coefficients of friction is considered in the design of the solution for each application.

[0050] German patent application DE 10 2021 129 423 A1 describes a ten-speed hub gear consisting of a five-speed input gearbox EGG and a secondary gearbox NSG, which includes a direct gear and a gear with a speed-increasing ratio. In direct gear, the secondary gearbox NSG rotates within its unit when the clutch KHR is open, and in the high-speed gear, the clutch KHR is closed. Due to the large gear ratio, approximately 35% of the input torque is transmitted to the hub sleeve, leaving 65% for the counter-torque to be absorbed by the main shaft. This counter-torque is shifted by the clutch KHR when the secondary gearbox NSG is shifted from high-speed to direct gear. The axially fixed part of the clutch KHR is rigidly connected to the ring gear HR5 via a side wall (see Fig. 3a) The axial limitation of the KHR coupling's range of motion is ensured by the central sleeve MH and the output-side flange of the main shaft HA. The axially movable part of the KHR coupling is directly and rotationally fixed to the main shaft HA via its splined connection and is controlled by a shift finger via the control groove NHR in the shift drum ST. To open the KHR coupling, both the static friction between the movable coupling part and the main shaft HA, as well as the static friction between the contacting surfaces of the coupling teeth, must be overcome.

[0051] The following describes exemplary embodiments of the invention as it is used within a complete device such as a coupling assembly or a system.

[0052] Preferably, the torque-dependent axial forces in the opening direction of the axial coupling are generated by a rotated position of the normals to the effective surfaces of the meshing coupling teeth. Advantageously, these normals form an angle β with the orthogonal to the axis of rotation 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 carried out by a smaller angular position of the normal, the control groove NHR, as in Fig. 3b shown, retained. With complete compensation or overcompensation by the helical gearing(s), the control groove NHR is maintained over the entire rotation angle of the shift drum, as shown in Fig. 3c shown, designed with a constant width adapted to the switching fingers so that the axial coupling remains securely closed in the closed state.

[0053] Another preferred embodiment of the present invention is described in the Fig. Figures 4a to 4c illustrate this. A load-operated axial clutch is a rotary block clutch that connects or disconnects two rotating parts. For example, this is clutch K70 of a ten-speed hub gear from German patent application DE 10 20 21 129 423 A1. The reference numerals used therein are also used and supplemented here.

[0054] The K70 coupling connects a sun gear SR3 to a central sleeve MH, which also serves as a partial web for the output-side part of an input gearbox EGG. The movable part of the axial coupling is guided axially and rotationally secured within the central sleeve MH by means of a splined connection. The axially fixed part of the K70 coupling is integrally integrated on the output side of the axially tightly guided sun gear SR3. The movable part of the K70 coupling has an internal annular collar that serves as a bearing surface for a sliding ring, which is guided axially and rotationally secured on the main axis HA.

[0055] The sliding ring is designed such that the collar inside the movable coupling part K70 rotates freely and is guided tightly by a rotationally secured, spring-loaded stop disc and an axially offset stop surface of the sliding ring. The axial length of the shoulder is only slightly greater than the wall thickness of the inner annular collar of the movable coupling part. This prevents a clamping effect. A spring F71 acting in the opening direction secures the open position of the coupling K70. A spring F70 acting in the closing direction assists the closing process of the coupling K70, as a track change occurs in a control groove N70 on a shift drum ST.

[0056] If static friction is only partially compensated, the sliding ring and thrust washer AS are installed rotated by 180 degrees, and spring F71 is omitted. This is because, now assisted by the axial forces, the clutch K70 opens against the spring force of spring F70. This applies equally to the other examples.

[0057] The K70 coupling allows for a combination of chamfered splines on both the plug-in and coupling teeth. Preferably, the coupling teeth are chamfered.

[0058] Another preferred embodiment of the present invention of a load-switchable axial coupling relates to a coupling that is located on the outer surface of a rotating sleeve arranged coaxially to a main axis HA and is actuated through this sleeve. An example of this embodiment is the coupling 90 from German patent application DE10 2018 007 326 A1. In contrast to the prior art, the control of the axial coupling, which is assisted by helix angles during disengagement, is advantageously carried out selectively via at least one actuator. The compensation of the frictional forces can preferably be distributed between the splined connection and / or the coupling teeth.

[0059] Another preferred embodiment of the present invention of a load-operated axial clutch relates to an axial clutch that functions as a locking device for a sun gear and must be engaged under load. Clutches of this type are 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 EGG input transmissions therein. This input transmission has, in the first sub-transmission, a complete planetary gear set with sun gear, planet gear, and ring gear, the drive of which is the planet carrier and the output via the ring gear. For gear changes, the sun gear is locked or released. A locking device is required for this purpose. When releasing the axially tightly guided sun gear, the axial clutch must be opened under load.

[0060] Within the sun gear, according to the invention, in this exemplary coupling assembly or this exemplary system or in this exemplary overall device, the movable coupling part is guided concentrically to it, rotationally fixed and axially displaceable by means of a splined connection, and to support the switching process, the splined connection preferably has a helix angle β S and the effective surfaces of the coupling teeth relative to an imaginary plane through the axis of rotation 210 by the angle β K twisted.

[0061] The rotationally fixed and axially fixed part of the axial coupling is preferably connected to the main shaft HA via a splined connection. The axial locking is achieved as shown in Fig. 5a shows a product made with screws that engage in holes adapted to the screw. Shown in Fig. 5c describes the control of the axial coupling in the case of compensation and overcompensation of frictional forces. The control of the axial coupling corresponds to that of coupling K70. The same applies here in the case of undercompensation.

[0062] The present invention comprises an axial coupling with a first shaft and with a second shaft which can be coupled and uncoupled controlled by an actuator, comprising a first working surface pairing WP1 with a first helix angle β1 and a second working 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 and the second helix angle β2 comprises an angular range of 0 degrees to 30 degrees and the sum of the helix angles β1 plus β2, via which a torque-dependent axial force AGES can be generated and which supports the uncoupling process of the axial coupling under load, comprises an angular range of 3 degrees to 35 degrees.

[0063] Axial couplings are preferably used in two-wheelers driven exclusively by pedals or manual drive, or also in two-wheelers additionally or exclusively driven by motors, especially in e-bikes or bicycles and cargo bikes assisted by electric motors. List of characters

[0064] The following section describes exemplary implementations and the derivation of the underlying dimensioning method, illustrated with figures. These figures show: Fig. 1a: an axial coupling in the prior art with straight-toothed plug-in teeth and orthogonal coupling teeth, Fig. 1b: an axial coupling with helical splined connection and the effective surfaces of the coupling teeth tilted relative to an imaginary plane through the axis of rotation, Fig. 1c: an orthogonal view in the direction of the axis of rotation with a section marker, Fig. 1d: the excerpt 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, Fig. 1e: a model of the coupling assembly made of Fig. 1b, supplemented by a force diagram for a torque which wants to rotate the sun wheel clockwise when viewed from the left, Fig. 2: the schematic gearbox section of a rear wheel hub from DE 10 2021 129 423 A1, Fig. 3a: the installation situation of a KHR coupling of a preferred embodiment of the invention, Fig. 3b: a preferred embodiment of a NHR control groove for the case of undercompensation, Fig. 3c: a further preferred embodiment of a control groove NHR for the case of compensation or overcompensation of axial forces, Fig. 3d: the installation situation of a KSR coupling, which is automatically switched indirectly by means of a KHR coupling via inclined running and / or plug-in teeth, Fig. 4a: a preferred embodiment of a load-switchable clutch K70, which is a block circulation clutch, Fig. 4b: 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, 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 K40 coupling in an isolated and more detailed representation with compensation or overcompensation of the axial forces, Fig. 5b: a preferred embodiment of a control groove N40 for the case of undercompensation, Fig. 5c: a preferred embodiment of a control groove N40 for the case of compensation or overcompensation, Fig. 6: a preferred embodiment of an installation situation of a K90 coupling and Fig. 7: a preferred embodiment of a section of an input gearbox EGG with helical gear teeth. 6. Detailed description

[0065] Identical components have the same reference symbols.

[0066] Fig. Figure 1a shows a prior art axial coupling with straight-toothed plug-in teeth and orthogonal coupling teeth.

[0067] Fig. Figure 1b shows an axial coupling with helical splined connection and the effective surfaces of the coupling teeth tilted relative to an imaginary plane through the axis of rotation.

[0068] A first shaft 200 transmits a torque M1 to the axial coupling 100. The first shaft 200 is either connected to a spur gear 120 via a toothed connection (not shown) or rigidly connected to it.

[0069] A second shaft 220, when the axial coupling 100 is coupled, takes over a torque M2, which is equal in magnitude to M1.

[0070] The first shaft 200 and the second shaft 220 are, in principle, rotatably mounted about a rotational axis 210 in a gear arrangement, 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.

[0071] The axial coupling 100 comprises a coupling ring 130 which, in the engaged state, engages a coupling toothing 114. This toothing is either rotationally fixed or, alternatively, integrally connected to a splined shaft 113. The splined shaft 113 is axially displaceable and rotationally fixed to the spur gear 120. The splined shaft 113 comprises at least one tooth 112, which has a helix angle β. S exhibits temperatures ranging from 0 degrees to 30 degrees.

[0072] The coupling ring 130 comprises at least one asymmetric tooth 131, which, in the engaged state, engages with a similarly asymmetric toothing 114 of the movable coupling part 110. The asymmetric teeth 131, 114 have a steep flank 132 and a shallower ramp 133 (see Fig. 1d). The ramp 133 has an inclination angle β3 in the range of 45 degrees to 75 degrees relative to an imaginary plane through the axis of rotation 210 (see Fig. 1d). This serves to reliably overcompensate for the frictional forces so that the clutch can operate as a freewheel. The flank 132 has a skew angle β relative to an imaginary plane through the axis of rotation 210. K in the range of 0 degrees to max. 30 degrees (see Fig. 1d).

[0073] The two inclined angles β S and β K or the two inclined angles β L and β K or the two inclined angles β L and β S each amount to a total of 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 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 moved by an actuator 140 (see Fig. 1e) is controlled in the opening direction x.

[0074] The sum of the two respective skew 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.

[0075] The axial force components A S and A K help the frictional forces R S and R K to overcome. This involves undercompensation, compensation, or overcompensation of the frictional forces R. S and R K generateable. The axial forces A S and A K are dependent on the difference between the magnitude of the input torque acting on the axial coupling 100 of the respective gear stage to be disengaged and the magnitude of the output torque of the respective gear stage to be disengaged.

[0076] The axial forces A S and A Ksupport the opening of the axial coupling 100 in the opening direction x, as soon as it is actuated by an actuator 140 (see Fig. 1e) is controlled in the opening direction x.

[0077] The preferred materials are steel and aluminum or steel and steel. However, other materials are also considered in combination with or with steel, 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 fibers, in order to achieve the smallest possible friction angle ρ. S and ρ K to be able to realize, in which safe engagement and disengagement between the movable coupling part 110 and the coupling ring 130 must be guaranteed even under high load and strong external vibrations.

[0078] The axial coupling comprises a material pairing between the coupling ring 130 and the coupling toothing 114 and between the plug toothing 113 and the guide of the spur gear 120 a material pairing of steel, aluminium, aluminium multi-component bronze, bronze, beryllium copper, brass, industrial ceramics or solid or surface-coated components carbon or carbon fibers to ensure secure engagement and disengagement between the lock and the coupling ring.

[0079] A force closing the coupling is exerted on the movable coupling part 110 via optional springs 111; 111' against a stationary housing part or against an axially rigid component (see Fig. 1e).

[0080] The movable coupling part 110 is guided axially by the spur gear 120. The spur gear 120 has internal splines with a helix angle β. S on.

[0081] Fig. Figure 1c shows an orthogonal view in the direction of the axis of rotation with a section marker.

[0082] 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.

[0083] On the movable coupling part 110 with splined connection, the helix angle β is shown using the example of a tooth 112 of the splined connection. S and using the example of a tooth of the coupling gear 114, the helix angle β K shown, which together form a derived axially movable wedge model, which is in Fig. 1e is shown in more detail.

[0084] Fig. 1e shows a model of the coupling assembly made of Fig. 1b, supplemented by a force diagram for a torque that wants to rotate the sun gear clockwise when viewed from the left. There, we find the double key, which represents an analogous problem to an axial coupling with a splined connection and a coupling tooth. In this analogy, the key is the movable part of the coupling. It has two helical contact surfaces, for which the mathematical derivation is analogous (i = S for the splined connection and i = K for the coupling tooth). All forces are considered in terms of their magnitude. The normal forces N act on the contact surfaces. i This results in frictional forces that act parallel to the tooth flanks: Ri=μi*Ni

[0085] The tangential component of the normal force, N i * cos(βi), must be due to the applied tangential force F i to be applied. Additionally, the frictional force also has a component R. i* sin(βi) in the direction of the applied tangential force, which it must also exert. Therefore: Fi=Ni*cos(βi)+Ri*sin(βi)

[0086] Using equations / 1 / and / 2 / , the following relation can be derived: Ni=Fi / (cos(βi)+μi*sin(βi))

[0087] The applied tangential force F i depends on the applied torque and the effective pitch circle radius of the gearing: Fi=M / ri

[0088] In the axial direction, one component of the normal force acts, A i = N i * sin(βi), as well as a component of the frictional force, R i * cos(βi) in the opposite direction. The resulting axial force X i This results in: Xi=Ni*sin(βi)−Ri / cos(βi).

[0089] Using equations / 1 / to / 5 / , the axial force X can be determined. i to Xi=M / ri*[sin(βi)−μi*cos(βi)] / [cos(βi)+μi*sin(βi)] determine. For the special case A i The result of = 0 is: βi=arctan(μi)

[0090] A helix angle that satisfies this relation is also called a friction angle ρ. ¡ named.

[0091] The total axial force acting on the moving part of the coupling is the sum of the axial forces of the plug and coupling teeth, i.e.: XGE = XS + XK

[0092] In the case of complete compensation of the frictional forces, the sum of the axial forces should be zero, i.e., X GES = 0. Therefore: [sin(βK)−μK*cos(βK)] / [cos(βK)+μK*sin(βK)]=−rK / rS*[sin(βS)−μS*cos(βS)] / [cos(βS)+μS*sin(βS)]

[0093] β K is therefore dependent on β S , where a solution can be found by: βK=2*arctan([a*b+((a2+1)*(b2+1))1 / 2−1] / [a+b]) with: a=µK b=-r K / r S *[sin(β S )-µ S *cos(β S )] / [cos(β S )+µ S *sin(β S )]

[0094] Empirically, a sufficiently good approximation could be determined for common skew angles, given by: βK=ρK−(rK / rS)*(βS−ρS)

[0095] The derivation is analogous for the embodiments according to the invention comprising an inclined positioning of a plug-in and a running tooth or a running and a coupling tooth.

[0096] The following describes examples of specific embodiments of the invention as they are used within a complete device such as a clutch assembly or a transmission system.

[0097] Fig. Figure 2 shows a rear wheel hub with a multi-speed transmission formed from planetary gear sets which engage in helical grooves N24 to N74, NSR, NHR in a shift drum by means of axially displaceable sliding rings with shift fingers, wherein the shift fingers are also guided in radial slots oriented parallel to the transmission axis of a hollow main shaft HA enclosing them and thereby different gears can be shifted by releasing radially extending clutch rings with axial teeth according to the invention, which are coupled by the sliding rings in a spring-loaded manner, according to the invention, wherein the multi-speed transmission arranged on the main shaft consists of a five-speed transmission as input transmission EGG, which consists of a two-stage and a subsequent single-stage planetary gear set, and a subsequent single-stage planetary gear set NSG, which engages in a high-speed transmission.which is arranged on a central sleeve MH and, in each case, connects the first five gears to a hub sleeve NH, or, with a fixed ring gear HR4, transmits these gears in a double-coupled manner to the hub sleeve NH, wherein 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 under load at any given time.

[0098] At least two of the planetary gear sets PR1, PR2 are identical and these have interconnected ring-shaped ring gears HR2-3, and the planet carriers PT2, PT3 of these two planetary gear sets are axially and radially supported against each other by an intermediate bearing ZL.

[0099] The sun gear SR5 of the downstream gearbox NSG is mounted on the center sleeve MH and axially displaceable and rotationally fixed to the hub sleeve NH, and can be coupled to the output-side planet carrier PT3 of the input gearbox EGG according to the invention with a sun gear coupling KSR, and the ring gear HR4 of the downstream gearbox NSG can be coupled close to the axis with the main shaft HA with a ring gear coupling KHR according to the invention.

[0100] Fig. 3a shows the installation situation of a KHR coupling of a preferred embodiment of the invention.

[0101] Fig. Figure 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.

[0102] Fig. Figure 3c shows a preferred embodiment of a control groove NHR for the case of compensation or overcompensation of axial forces.

[0103] In Fig. Figure 3d shows a further simplified embodiment of the secondary transmission, in which the engagement of the sun gear clutch KSR is not effected by the shift drum ST, but rather the thrust forces of a helical gear in the secondary transmission and / or the helical gearing of the splined connection of the torque transducer in the hub sleeve NH are used in such a way that the sun gear clutch KSR is always automatically opened or closed against a spring force, in the opposite direction to the actively engaged ring gear clutch KHR. Accordingly, in direct drive, the sun gear clutch KSR is closed and the ring gear clutch KHR is open. The secondary transmission therefore rotates as a single unit.In the preferred variant, the thread chamfer and / or the splined connection chamfer is oriented such that, under load, the sun gear SR5 moves away from the stationary part of the coupling KSR due to the axial force, and the ring gear HR5 experiences an opposing force, which is supported against the hub sleeve NH by the axial support bearing ASL. The torque transmitter D5 is rotationally fixed but axially displaceable to the hub sleeve NH and rotationally and axially fixed to the sun gear SR5. When stationary or operating in direct drive of the downstream transmission NSG, the sun gear coupling KSR is closed by means of the return spring F5, which axially supports the torque transmitter D5 against the hub sleeve NH. This movement is limited by a stop in the hub sleeve NH.The axial forces of the ring gear HR5 are absorbed directly via an axially load-bearing rolling bearing between the hub sleeve NH and the ring gear HR5, and / or via an axial support element ASE on the main shaft HA. The shifting process is supported by the fact that, when the ring gear clutch KHR closes, the sun gear SR5 overtakes the planet carrier PT5 and thus switches to freewheel mode, with the orientation of the clutch teeth additionally ensuring that the clutch KSR opens. The drive of the hub sleeve NH is transferred from the center sleeve MH to the sun gear SR5 without interruption of the load.

[0104] Fig. Figure 4a shows a preferred embodiment of a clutch K70 according to the invention, which can be switched under load. It is a closed-loop clutch that connects or disconnects two rotating parts by switching. For example, this is 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. The reference numerals used therein are also used here and supplemented. The clutch K70 designed according to the invention in this system example connects the sun gear SR3 to the central 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 inside the central sleeve by means of a splined connection, preventing rotation. The axially fixed part of the clutch is integrally housed on the output side of the axially tightly guided sun gear SR3.The movable part of the coupling has an internal annular collar that serves as a bearing surface for the sliding ring, which is guided axially on the main axis and secured against rotation. The sliding ring is designed such that the collar inside the movable coupling part rotates freely and is tightly guided by a rotationally secured, spring-loaded stop ring and an axially offset stop surface of the sliding ring. The axial length of the shoulder is only slightly greater than the wall thickness of the internal annular collar of the movable coupling part. This prevents a clamping effect. The spring F71, acting in the opening direction, secures the open position of the coupling. The spring F70, acting in the closing direction, assists the closing process of the coupling, as a track change occurs in the control groove N70 on the shift drum ST.If static friction is only partially compensated, the sliding ring and thrust washer are installed rotated by 180 degrees, and spring F71 is omitted. This is because, now assisted by the axial forces, the clutch K70 opens against the spring force of F70. This applies equally to the other examples.

[0105] With the K70, a combination of helical threads for the splined connection and the coupling teeth can be used. Preferably, the coupling teeth are helical.

[0106] The axial forces A S and A K (see each) Fig. 1e) support the opening of the axial coupling 100 in the opening direction x (also in Fig. 1e shown), as soon as it is controlled by an actuator 140.

[0107] Fig. Figure 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.

[0108] Fig. Figure 4c shows a preferred embodiment of a control groove N70 for the case of compensation or overcompensation. The switching finger SF is closely guided by both flanks of the control groove N70.

[0109] Fig. Figure 5a shows a preferred embodiment of an installation situation of a coupling K40 according to the invention in an isolated and more detailed view with compensation or overcompensation of the axial forces. In this embodiment, a helical splined connection and / or a helical coupling connection is used.

[0110] It is a planetary gear stage whose drive is a planet carrier and whose output is via a ring gear. To change gears, a sun gear SR2 is either locked or unlocked. This requires a locking mechanism. When unlocking the sun gear SR2, the clutch K40 must be opened under load. Inside the sun gear SR2, the movable clutch element is guided concentrically to it by means of a splined connection SST, which prevents rotation but allows axial movement. To assist the shifting process, the splined connection SST has a helix angle β. S and / or the effective surfaces of the coupling teeth with respect to an imaginary plane through the axis of rotation of the gearbox by the angle β K twisted. The rotationally fixed and axially fixed part of the K40 coupling is connected to the main shaft HA via a splined connection SST. The axial locking is preferably achieved as shown in Fig. 5a shows that it is made using screws that engage in holes adapted to the screw. Shown in Fig. 5c describes the control of the axial coupling in the case of compensation and overcompensation of frictional forces. The control of coupling K40 corresponds to that of coupling K70. The same applies here in the case of undercompensation.

[0111] The axial forces A S and A K (see each) Fig. 1e) support the opening of the axial coupling 100 in the opening direction x (also in Fig. 1e shown), as soon as it is controlled by an actuator 140.

[0112] Fig. Figure 5b shows a preferred embodiment of a groove N40 for the case of undercompensation.

[0113] Fig. Figure 5c shows a preferred embodiment of a groove N40 for the case of compensation or overcompensation.

[0114] Fig. Figure 6 shows a preferred embodiment of an installation situation for a K90 coupling.

[0115] A preferred embodiment of the present invention for such a transmission is described in Fig. Figure 6 is shown schematically. The output-side clutch K90 of the NSG auxiliary transmission shown there is actuated via two cascaded shift fingers. One of these is guided in a groove N94 in the shift drum ST, and the other passes through the rotating output sleeve H and actuates a clutch ring 130. Two springs F90 and F91 center the stacked shift fingers. When the clutch K90 is engaged, the output torque is transmitted from the output-side side wall of the ring gear HR to the output sleeve, which can rotate freely because a retaining ring SR fixes its axial play on the output sleeve H.

[0116] According to another preferred embodiment, an axial coupling 100 is positioned closer to a main axis. The axial coupling 100 is located in a rotating sleeve H. Helical gearing, used to facilitate switching, is arranged between a planet carrier PT and an output-side coupling half. Switching fingers SF and centering springs also engage this half. Stacking of switching fingers SF is therefore eliminated, which significantly simplifies the design and manufacturing.

[0117] In the clutches of the transmission according to DE 10 2018 007 A1, the switching forces of some clutches K50 and K90 in the output-side transmission block are fed to the shift rings via several stages, whereby rotating sleeves of shift elements are inserted. This can result in axial forces at the shift elements that are difficult to control or are necessary for shifting if the transmission gear teeth are misaligned. Therefore, only slight misalignments of a few degrees are permitted there.

[0118] The axial forces A S and A K (see each) Fig. 1e) support the opening of the axial coupling 100 in the opening direction x (also in Fig. 1e shown), as soon as it is controlled by an actuator 140.

[0119] Fig. Figure 7 shows a section of the input gearbox EGG with at least one axial coupling 100 according to the invention. The two symmetrical planetary gear sets with planet gears PR2 and PR3 are provided with helical gears, which are shown symbolically. 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 couplings K40, K50 according to the invention. The reaction forces acting on the interconnected ring gears HR2 and HR3 via the planet gears PR2, PR3 compensate each other and hold them centered on each other as a so-called flying ring pair. The opening forces A S and A KIn the opposite direction are closing springs F40, F50. This means that relatively moderate switching forces are required at the shift fingers, which each engage in a loose fit in control grooves N40, N50, in order to accomplish the switching of the axial clutches even under high transmission load.

[0120] The release springs F41, F51 are dimensioned so that the sliding rings secure open clutch positions, even when no auxiliary force is available from the helical gearing due to lack of or low gearbox load.

[0121] It is advantageous that the sun gears and the associated clutch ring can be manufactured as a single piece. In the described rear-wheel transmission, a clutch travel of only approximately 1.5 mm between the closed and open positions is sufficient.

[0122] At Fig. 7. The axial forces of the running gear and the clutch gear are utilized. Only clutch K40 is opened under load. Clutch K50 enters freewheel mode when clutch K70, the rotary block clutch, is closed. The sun gear SR3 is then simply driven in the direction of rotation. Therefore, the chamfer on clutch K50 could be omitted. The control remains as it is in the Fig. Figure 2 shows that this provides a design that makes it possible to find good solutions even in difficult space conditions. In this case, the axial forces of the K50 coupling are undercompensated.

[0123] The K40 clutch acts as a forward-running lock and the K50 clutch acts as a reverse-running lock.

[0124] In a preferred embodiment, the helical gearing is located on both the running gear teeth and the clutch gear teeth. This is due to the geometric and tribological situation, since only clutch K40 needs to be engaged under load. Assuming a rolling friction coefficient of µ for the running gear teeth... L = 0.015, which corresponds to a friction angle ρ L of 0.86 degrees corresponds, relative to a partial circle radius r L of 25 mm, a mean pitch circle radius r K the coupling teeth of 17.2 mm with a coefficient of friction µ K of 0.12, which corresponds to a friction angle ρ K If the angle of 6.84 degrees corresponds to the following design variants when compensating for frictional forces according to the inventive method: βK=0=>βL=10.9 degrees βK=ρK=6.84 degrees,βL=ρL=0.86 degrees, βK=3 degrees=>βL=6.44 degrees as well as βK=5 degrees=>βL=3.54 degrees.

[0125] It can be seen that the sum of the friction angles ρ K and ρ L only in the case of individually compensated gear teeth that the actually necessary angle sum corresponds.

[0126] The axial force AGES assists in opening the axial coupling 100 in the opening direction x, as soon as it is actuated by an actuator 140 (see Fig. 1e) is controlled in the opening direction x.

[0127] The Fig. 2 to Fig.Figure 7 shows 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 is subjected to a closing force by a closing spring 111, 111' and also engages with at least one actuator 140, comprising at least one switching finger, in a partially helical guide groove of a switching drum, so that in predetermined rotational positions of the same, the side wall of the groove exerts a switching force against the closing force on the switching fingers and thereby opens the axial clutch 100 and thus establishes a different switching state of the respective gear stage, characterized in that the toothing of the planetary gear is designed as a helical toothing,that, as a result, torque-dependent axial forces AGES occurring under load act 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 on the coupling elements are absorbed by support elements on adjacent gearbox components. The reference list refers to the attached figures. 100 axial coupling 110 movable coupling part 111; 111' Feather 112 at least one tooth of the splined connection 113 Splined connection 114 Clutch teeth 120 Spur gear 130 Clutch ring 131 asymmetrical teeth 132 steep flank of the coupling teeth 133 flat ramp of the coupling teeth 140 actuator 200 first wave 220 second wave 210 Rotation axis 300 sliding ring with collar 600 gear teeth WP1 first effective surface pairing WP2 second effective surface pairing MP1 first material pairing MP2 second material pairing ps Friction angle Plug-in spline ρ K Friction angle of clutch teeth P L Friction angle of the gear teeth β S Inclination angle of 113 β K Inclination angle of 114 β L Bevel angle of 600 β1 first inclination angle β2 second inclined angle β3 Inclination angle of 133 µ S Friction coefficient of splined connection µ K Friction coefficient of clutch teeth µ L Friction coefficient of the gear teeth r S Pitch circle radius of splined shaft r K mean pitch circle radius coupling teeth r L Pitch circle radius of the gear teeth N S , N K Normal force plug / coupling toothing R S , R K Frictional force of plug / coupling teeth F S , F K Tangential force plug / coupling toothing A S , A K axial component of N S , N K A GES Sum of the axial forces generated by the skew of gears X S , X K Effective axial force at plug / coupling toothing X GES Sum of effective axial forces x, opening direction M1, M2 torque M magnitude of torque M1 / M2 AH drive sleeve HA Main Axle MH Middle sleeve NH hub sleeve ST shift drum N20, N30, N40, N50, N60, N70, N80, N90, NSR, NHR Taxpayer(s) L Lager von MH PT1-2, PT3 / PT3-4, PT5 planet carrier ZL Intermediate storage PT1-2 to PT3-4 SR1 - SR5 sun gears PR1 - PR5 planetary gears HR2-3, HR5 hollow gears F5 Return spring F40, F50 closing spring F41, F51 Release spring F70, F71 spring ASL axial support bearing ASE axial support element D5 torque transmitter KR rest positions AS thrust washer IGH Integrated Gear Hub HR ring gear PR planetary gear SR sun wheel EGG input gearbox NSG secondary gearbox G1 ... G10 switching positions K20, K30, K40 clutches first sub-transmission K50, (K60), K70, K80 clutches second sub-transmission KSR sun gear coupling in the NSG KHR hollow gear coupling in the NSG SST helical splined connection SF shift finger K backdrop H sleeve

Claims

[1] Axial coupling (100) with a first shaft (200) and with 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 generated and which supports the uncoupling process of the axial coupling (100) under load, encompasses an angular range of 3 degrees to 35 degrees, wherein the first effective surface pairing (WP1) and the second effective surface pairing (WP2) are each formed by a toothing, the toothings being a coupling toothing and a running toothing of spur gears of a planetary gear or a spur gear, or a plug toothing and a running toothing of spur gears of a planetary gear or a spur gear, and the torque-dependent axial force (A) is generated by the inclination of the toothings under drive load. GES ) can be generated, which acts in the opening direction (x) of the axial coupling (100) and supports its uncoupling process under load. [2] Axial coupling (100) according to claim 1 wherein the first contact surface pairing (WP1) comprises the 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 contact surface pairing (WP2) includes the running gear (600) and the second helix angle (β2) a helix angle (β L ) includes [3] Axial coupling (100) according to claim 1 wherein the first contact surface pairing (WP1) comprises the coupling toothing (114) with at least one tooth (112), and the first helix angle (β1) has a helix angle (β K ) includes and the second contact surface pairing (WP2) includes the running gear (600) and the second helix angle (β2) a helix angle (β L ) includes [4] Axial coupling (100) according to one of the preceding claims, wherein the first material pairing (MP1) of the first working surface pairing (WP1) and the second material pairing (MP2) of the second working surface pairing (WP2) comprise one of the material pairings of steel, bronze, beryllium copper, aluminium, aluminium multi-component bronze, ceramic, carbon or carbon fibers individually or in combination to ensure safe uncoupling. [5] 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 on the other hand 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 the side wall of the groove exerts a switching force against the closing force on the switching fingers and thereby opens the axial clutch (100) and thus establishes a different switching state of the respective gear stage, characterized by , that the gearing of the planetary gear is designed as a helical gear in such a way 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 on the coupling elements are absorbed by support elements on adjacent transmission components. [6] Coupling assembly comprising a planetary gear unit according to claim 5, characterized by, that sun gears (SR2, SR3) of two symmetrically adjacent helical planetary gear sets are connected to and disconnected from a main shaft (HA) by couplings (K40, K50), whereby the reaction forces of the sun gears that occur, 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 connection. [7] Coupling assembly comprising a planetary gear unit according to claim 8, characterized by that the clutch rings are each connected to the clutch (K40, K50). [8] Coupling assembly comprising a planetary gear unit according to any one of claims 5 to 7, characterized by that helical gearing is a running gear whose helix angle is between 5 degrees and 12 degrees, in particular between 8 degrees and 10 degrees. [9] Clutch assembly comprising a planetary gear unit according to claim 5, characterized by that the helical gearing is a splined gearing and the ring gear is supported by a support bearing on the hub sleeve and above that on the gearbox housing. [10] Coupling assembly comprising a planetary gear unit according to claim 5, characterized by , that the shift finger, which is guided in the groove (N60 / N80) with low tolerance in the control drum (ST), is held axially centered on both sides of the main shaft (HA) between two springs, each of which is axially fixed at its end by support discs on the main shaft, drives an axially displaceable sun gear 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 to be shifted (K60, K80).

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