DE102022107915B3 - Improved shifting of flat gear axial clutches under load - Google Patents

Abstract

The present invention relates to an axial clutch with a first shaft and with a second shaft, which can be coupled and uncoupled in a controlled manner by an actuator, comprising a first pair of active surfaces with a first helix angle and a second pair of active surfaces with a second helix angle, the first helix angle having an angular range of 0 degrees to 30 degrees relative to an imaginary plane through an axis of rotation and the second helix angle comprises an angle range of 0 degrees to 30 degrees relative to an imaginary plane through an axis of rotation and the sum of the helix angles over which a torque-dependent axial force can be generated and which the uncoupling process of the axial clutch supported under load, covers an angle range of 3 degrees to 35 degrees. Furthermore, the present invention relates to a method for dimensioning an axial clutch.

Description

1. Technical field

The present invention relates to an arrangement of planetary gears or spur gears that can be shifted into different gears, which is provided with at least one axial clutch, which is optionally kept closed by the pressing force of a spring on an axially toothed clutch ring, which engages with control fingers in a control groove of a control cylinder the rotation of which, for the purpose of shifting gears, engages or disengages the clutch in question by means of a shifting force which occurs on the side wall of the groove and is directed against the pressing force.

2. State of the art

Such gears with axial couplings for the purpose of setting different gears are, for example, planetary gears from the German patent application DE 10 2018 007 326 A1 previously known and in the German patent application DE 10 2021 129 423 A1 described. All of the gears in the transmissions mentioned can also be switched under an operating load. However, with individual shifting 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 operation and/or the auxiliary motor is indicated for the corresponding shifting operations. In the DE 10 2021 129 423 A1 a rear-mounted transmission is described which has helical running gearing which is used for indirect switching of the sun wheel clutch KSR. The shifting process is triggered by the change in status of the clutch KHR. The sun wheel clutch does not have an actuator which is controlled via a control groove in a switching drum. Furthermore, the clutch KSR is in freewheeling mode after the ring gear clutch KHR is closed, since the sun gear SR rotates faster than the carrier.

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

From the DE 11 2017 007 405 T5 a clutch mechanism for a bicycle or a bicycle with an additional electric drive is known.

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

A switching device and a vehicle drive train and a method for operating a switching device and a vehicle drive train is from DE 10 2013 213 156 A1 known.

3. Task

It is the object of the present invention to improve the shiftability of axial clutches of planetary gears or spur gears in such a way that they can also be securely coupled under drive load and can be disengaged by an actuator with low shifting forces.

4. Summary of the Invention

The object is solved according to the features of independent claim 1 and independent claim 14 . Further advantageous configurations of the solution can be found in the respective dependent claims.

1. a) eine Kupplungsverzahnung und eine Steckverzahnung oder
2. b) eine Kupplungsverzahnung und eine Laufverzahnung der Stirnräder oder
3. c) eine Steckverzahnung und eine Laufverzahnung der Stirnräder

eines Planetengetriebes oder eines Stirnradgetriebes hinsichtlich der Schrägstellung der Verzahnung so ausgebildet ist/sind, dass die unter Antriebslast auftretenden drehmomentabhängigen Axialkräfte in Öffnungsrichtung einer Axialkupplung wirken, und die durch die zwischen den sich berührenden Kupplungszähnen und den sich berührenden Steckverbindungszähnen oder zwischen den sich berührenden Kupplungszähnen und den sich berührenden Laufradzähnen, oder zwischen den sich berührenden Steckverbindungszähnen und den sich berührenden Laufradzähnen einer dem Entkuppeln der Axialkupplung entgegen gerichtet wirkende Haftreibung weitgehend kompensieren, wobei daraus resultierende an den Kupplungselementen auftretende Reaktionskräfte in Stützelementen an benachbarten Getriebekomponenten abgefangen werden.The solution is that the gearing, in particular

1. a) a clutch spline and a spline or
2. b) a clutch toothing and a running toothing of the spur gears or
3. c) a spline and a running gear of the spur gears

of a planetary gear or a spur gear is/are designed with regard to the inclined position 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 clutch, and the forces generated between the contacting clutch teeth and the contacting connector teeth or between the contacting clutch teeth and the contacting impeller teeth, or between the contacting connector teeth and the contacting impeller teeth, largely compensate for static friction acting in the opposite direction to the disengagement of the axial clutch, with the resultant reaction forces occurring on the clutch elements being absorbed in support elements on adjacent transmission components.

The spur gear or the planetary gear is arranged in a housing which is connected fixedly or by means of a torque arm connected to an axle in a rotationally and axially fixed manner to the frame of a vehicle which is driven via at least one wheel and can be moved over a roadway. The frame establishes a directional reference to the roadway. Typically, at least two wheels provide the directional reference between the frame and the roadway. In the case of only one drive wheel with no other wheels, the directional reference can be established by the driver (unicycle) or by a controller (segway). The starting point here is preferably a two-wheeler, which is driven by at least one of the wheels.

The gearbox 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.

A shaft within the meaning of this invention is a component that can absorb or transmit torque.

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

The incoming shaft and the outgoing shaft are rotatably mounted in the gear housing of the spur gear and usually in the gear housing of the planetary gear. In the case of the planetary gear, one of the shafts can be connected to the gear housing or the main axis in a rotationally fixed manner.

1. a) einer Kupplungsverzahnung mit einer derart um einen Schrägungswinkel β_K zur Rotationsachse hin gekippten Flächennormalen der Wirkflächen, dass der Winkel zwischen Wirkfläche und Rückflanke der Kupplungszähne vergrößert wird oder
2. b) einer schrägverzahnten Laufverzahnung mit einem Schrägungswinkel β_L oder
3. c) einer schrägverzahnten Steckverzahnung mit einem Schrägungswinkel β_S erzeugt.

Torque-dependent axial forces of

1. a) a clutch toothing with a surface normal of the active surfaces tilted by a helix angle β _K to the axis of rotation such that the angle between the active surface and the trailing edge of the clutch teeth is increased or
2. b) a helical running gear with a helix angle β _L or
3. c) a helical spline with a helix angle β _S generated.

If the respective helix angle β _K or β _L or β _S is equal to the arctangent of the coefficient of friction μ of the pairs of opposing active surfaces in the respective material pairing, the axial force corresponds in terms of amount to the axial component of the frictional force. This angle is referred to below as the angle of friction p.

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

If the axially movable part of an axial clutch is given the degree of freedom necessary for engaging in the axially fixed clutch part by means of a spline, the axial mobility of the component carrying the clutch teeth being increased by the axial displaceability inherent in a spline of the component with the clutch toothing in relation to an axially closely guided, driving component is achieved, the axial clutch can be opened and closed by means of the actuator provided. This allows potential rotation of the coupled shaft through the flat-toothed axial coupling. The axial force required to open the axial clutch to overcome the friction between the effective surfaces of the spline and the effective surfaces of the flat-toothed clutch splines 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 on the material pairing of the effective surface pair and the mean pitch circle diameter the clutch teeth. Below the middle pitch circle radius the clutch teeth. is to be understood as the mean radius of the circular ring that completely encloses the radially aligned clutch teeth.

If the axially movable part of an axial clutch is given the degree of freedom necessary for engaging in the axially fixed clutch part by means of a running gear, the axial mobility of the gear wheel carrying the clutch teeth being increased by the axial displaceability of the gear wheel with the clutch gearing inherent in running gearing compared to an axially closely guided, driving gear wheel is achieved, the axial clutch can be opened and closed by means of the actuator provided. This allows potential rotation of the coupled shaft through the flat-toothed axial coupling. The axial force required to open the axial clutch to overcome the friction between the effective surfaces of the running gear and the effective surfaces of the flat-toothed clutch gearing 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 on the material pairing of the active surface pair and the mean pitch circle diameter the clutch teeth.

If the axially movable part of an axial clutch is given the degree of freedom necessary for engaging in the axially fixed clutch part by means of a running gear and a spline, the axial mobility of the gear wheel carrying the clutch teeth being increased by the axial displaceability of the gear wheel with the clutch gearing and the Spline is achieved in relation to an axially closely guided, driving gear, the axial clutch, which is opened indirectly via an actuator when opening in the freewheel mode and closed by means of a return spring when the drive torque of the driving gear is lost. This allows potential rotation of the disengaged shaft relative to the previously coupled shaft.

The axial force required to open the axial clutch 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 on the material pairing of the active surface pair and the pitch circle diameter of the spline away. Since the axial clutch switches to freewheeling mode when it is opened, there are no frictional forces between the active surfaces of the clutch teeth. With a clutch of this type, the drive shaft is changed under load.

The following applies: the smaller the respective radius, the greater the frictional force at a given torque.

In the variants according to the invention there are in each case two pairs of active surfaces, on which a first and a second frictional force is generated, the friction values of which depend on a first and a second pair of materials.

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

Some reference symbols are inserted below for better understanding, which are also used in the detailed description of the figures and the derivation of the calculation method.

• • Bestimmung des Reibungskoeffizienten µ₁ zwischen dem an einer ersten Materialpaarung MP1 beteiligten Wirkflächenpaar WP1 und des Reibungskoeffizienten µ₂ zwischen dem an einer zweiten Materialpaarung MP2 beteiligten Wirkflächenpaar WP2 mittels üblicher Verfahren am Objekt und/oder aus einer Tabelle mit Materialeigenschaften.
• • Bestimmung des wirksamen Teilkreisradius r₁ der ersten Wirkflächenpaarung WP1 und des wirksamen Teilkreisradius r₂ der zweiten Wirkflächenpaarung WP2 aus den Konstruktionsdaten.
• • Bestimmung oder Festlegung eines ersten Schrägungswinkels β₁ der ersten Wirkflächenpaarung WP1 und eines zweiten Schrägungswinkels β₂ der zweiten Wirkflächenpaarung WP2 aus solchen Konstruktionsdaten, die lastunabhängig sind, nämlich den Reibungskoeffizienten der ersten Materialpaarung MP1 und der zweiten Materialpaarung MP2 µ₁ und µ₂, den Teilkreisradien der ersten Wirkflächenpaarung WP1 und der zweiten Wirkflächenpaarung WP2 r₁ und r₂, wobei sich z. B. im Falle vollständiger Kompensation der Reibkräfte der Winkel β₂ als Funktion von β₁ bestimmen lässt:

$${\mathrm{\beta }}_2=2*\text{arctan}\left(\left[\text{a*b+}{\left(\left({\text{a}}^2+1\right)*\left({\text{b}}^2+1\right)\right)}^{1/2}-1\right]/\left[\text{a}+\text{b}\right]\right)$$

mit: $$\text{a}={\mathrm{\mu }}_2$$

$$\text{b}=-{\text{r}}_2/{\text{<figure-callout id="1" label="r" filenames="DE102022107915B3\_0035.png" state="{{state}}">r</figure-callout>}}_1*\left[\text{sin}\left({\mathrm{\beta }}_1\right)-{\mathrm{<figure-callout\; id="1"\; label="\mu "\; filenames="DE102022107915B3\_0035.png"\; state="\{\{state\}\}">\mu </figure-callout>}}_1*\text{cos}\left({\mathrm{\beta }}_1\right)\right]/\left[\text{cos}\left({\mathrm{\beta }}_1\right)+{\mathrm{<figure-callout\; id="1"\; label="\mu "\; filenames="DE102022107915B3\_0035.png"\; state="\{\{state\}\}">\mu </figure-callout>}}_1*\text{sin}\left({\mathrm{\beta }}_1\right)\right]$$

The present invention also solves the problem by means of a method for dimensioning the axial clutches according to the invention, taking into account the geometric and tribological conditions, with the following steps:

• • Determination of the coefficient of friction μ ₁ between the active surface pair WP1 involved in a first material pairing MP1 and the friction coefficient μ ₂ between the active surface pair WP2 involved in a second material pairing MP2 using conventional methods on the object and/or from a table with material properties.
• • Determination of the effective pitch circle radius r ₁ of the first effective surface pairing WP1 and the effective pitch circle radius r ₂ of the second effective surface pairing WP2 from the design data.
• • Determination or specification of a first helix angle β ₁ of the first pair of active surfaces WP1 and a second helix angle β ₂ of the second pair of active surfaces WP2 from such design data that are load-independent, namely the coefficient of friction of the first pair of materials MP1 and the second pair of materials MP2 µ ₁ and µ ₂ , the Pitch circle radii of the first effective surface pairs tion WP1 and the second effective surface pairing WP2 r ₁ and r ₂ , where z. B. in the case of complete compensation of the frictional forces, the angle β ₂ can be determined as a function of β ₁ :

$${\mathrm{\beta }}_2=2*\text{arctan}\left(\left[\text{a*b+}{\left(\left({\text{a}}^2+1\right)*\left({\text{b}}^2+1\right)\right)}^{1/2}-1\right]/\left[\text{a}+\text{b}\right]\right)$$

With: $$\text{a}={\mathrm{\mu }}_2$$

$$\text{b}=-{\text{right}}_2/{\text{right}}_1*\left[\text{sin}\left({\mathrm{\beta }}_1\right)-{\mathrm{<figure-callout\; id="1"\; label="\mu "\; filenames="DE102022107915B3\_0035.png"\; state="\{\{state\}\}">\mu </figure-callout>}}_1*\text{cos}\left({\mathrm{\beta }}_1\right)\right]/\left[\text{cos}\left({\mathrm{\beta }}_1\right)+{\mathrm{<figure-callout\; id="1"\; label="\mu "\; filenames="DE102022107915B3\_0035.png"\; state="\{\{state\}\}">\mu </figure-callout>}}_1*\text{sin}\left({\mathrm{\beta }}_1\right)\right]$$

mit: $${\mathrm{\rho }}_{\text{i}}=\text{arctan}\left({\mathrm{\mu }}_{\text{i}}\right),\text{ i}=\mathrm{1,2}$$

i = 1,2A sufficiently good approximation for angles up to 30° results from: $${\mathrm{\beta }}_2={\mathrm{\rho }}_2-\left({\text{right}}_2/{\text{right}}_1\right)*\left({\mathrm{<figure-callout\; id="1"\; label="\beta "\; filenames="DE102022107915B3\_0035.png"\; state="\{\{state\}\}">\beta </figure-callout>}}_1-{\mathrm{\rho }}_1\right)$$

With: $${\mathrm{\rho }}_{\text{i}}=\text{arctan}\left({\mathrm{\mu }}_{\text{i}}\right),\text{i}=1.2$$

i = 1.2

A simple solution is to compensate for the frictional forces individually, ie: $${\mathrm{\beta }}_2={\mathrm{\rho }}_2\text{and}{\mathrm{<figure-callout\; id="1"\; label="\beta "\; filenames="DE102022107915B3\_0035.png"\; state="\{\{state\}\}">\beta </figure-callout>}}_1={\mathrm{<figure-callout\; id="1"\; label="\rho "\; filenames="DE102022107915B3\_0035.png"\; state="\{\{state\}\}">\rho </figure-callout>}}_1.$$

In a first preferred and alternative embodiment of the axial clutch, this includes to support the uncoupling process of the axial clutch in a first pair of active surfaces WP1 with material pairing MP1 splines and in a second pair of active surfaces WP2 with material pairing MP2 clutch teeth.

In this embodiment, the helix angle β ₁ is formed by a helix angle β _S of the spline. The helix angle β ₂ is formed by a helix angle β _K of the clutch teeth.

The pitch circle radius r ₁ forms a pitch circle radius rs of the spline. The pitch circle radius r ₂ forms a mean pitch circle radius r _K of the clutch teeth.

The coefficient of friction μ ₁ results from the first material pairing MP1 of the spline with a coefficient of friction μ _S . The coefficient of friction μ ₂ results from the second material pairing MP2 of the clutch teeth with a coefficient of friction μ _K .

In a further alternative embodiment of the axial clutch, this comprises a spline in the first active surface pairing WP1 with material pairing MP1 and a running gear in the second active surface pairing WP2 with material pairing MP2 to support the uncoupling process of the axial clutch.

In this embodiment, the helix angle β ₁ is formed by a helix angle β _S of the spline. The helix angle β ₂ is formed by a helix angle β _L of the running teeth.

The pitch circle radius r ₁ forms a pitch circle radius r _s of the spline. The pitch circle radius r ₂ forms a pitch circle radius r _L of the running teeth.

The coefficient of friction μ ₁ results from the first material pairing MP1 of the spline with a coefficient of friction μ _S . The coefficient of friction μ ₂ results from the second material pairing MP2 of the running teeth with a coefficient of friction μ _L .

In a further alternative embodiment of the axial clutch, this comprises clutch teeth in the first active surface pairing WP1 with material pairing MP1 and running gearing in the second active surface pairing WP2 with material pairing MP2 to support the uncoupling process of the axial clutch.

In this embodiment, the helix angle β ₁ is formed by a helix angle β _K of the clutch teeth. The helix angle β ₂ is formed by a helix angle β _L of the running teeth.

The pitch circle radius r ₁ forms a mean pitch circle radius r _K of the clutch teeth. The pitch circle radius r ₂ forms a pitch circle radius r _L of the running teeth.

The coefficient of friction μ ₁ results from the first material pairing MP1 of the clutch teeth with a coefficient of friction μ _k . The coefficient of friction μ ₂ results from the second material pairing MP2 of the running teeth with a coefficient of friction μL.

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

With axial clutch 100, a first shaft 200 can be coupled and uncoupled with a second shaft 220, controlled by an actuator 140, comprising a first pair of active surfaces WP1 with a first helix angle β ₁ and a second pair of active surfaces WP2 with a second helix angle β ₂ , the first helix angle β ₁ comprises an angular range of 0 degrees to 30 degrees relative to an imaginary plane through an axis of rotation 210 and the second helix angle β ₂ comprises an angular range of 0 degrees to 30 degrees relative to an imaginary plane through an axis of rotation 210 and the sum of the helix angles β ₁ plus β ₂ , via which a torque-dependent axial force AGES can be generated and which supports the uncoupling process of the axial clutch 100 under load, includes an angular range of 3 degrees to 35 degrees.

In a first embodiment, the axial clutch 100 comprises in the first pair of active surfaces WP1 a spline 113 with at least one tooth 112 and the first helix angle β ₁ comprises a helix angle β _S and in the second pair of active surfaces WP2 clutch teeth 114 and the second helix angle β ₂ comprises a helix angle β _K .

A coupling ring 130 of the coupling toothing 114 can be provided, which comprises an asymmetrical toothing 131 with at least one tooth 112, further comprising a flank 132 which has the angle β _K .

The coupling ring 130 is designed to be controllable by an actuator 140 in an opening direction x, with the axial force A _GES supporting the opening of the axial coupling 100 in the opening direction x.

In a further embodiment, the axial clutch 100 comprises in the first pair of active surfaces WP1 a spline 113 with at least one tooth 112 and the first helix angle β ₁ comprises a helix angle β _S and in the second pair of active surfaces WP2 a running gear 600 and the second helix angle β ₂ comprises a helix angle β _L .

In a further embodiment, the axial clutch 100 comprises a clutch toothing 114 with at least one tooth 112 in the first pair of active surfaces WP1 and the first helix angle β ₁ comprises a helix angle β _K and the second pair of active surfaces WP2 comprises a running tooth system 600 and the second helix angle β ₂ comprises a helix angle β _L .

The first material pairing MP1 and the second material pairing MP2 of axial clutch 100 can each be one of the material pairings of steel, bronze, beryllium copper, aluminum, aluminum multi-component 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 engagement and disengagement.

Preferred embodiments are explained in detail in the following examples. It is taken into account here that the coefficients of friction depend on the pairing of materials, the heat treatment such as hardening, lubrication, the duration of contact, etc. This means that a considerable scattering of the coefficients of friction is taken into account when designing the solution for each application.

In the German patent application DE 10 2021 129 423 A1 a ten-speed hub transmission is described, which consists of a five-speed input transmission EGG and a rear-mounted transmission NSG, which has a direct gear and one with a speed-up gear ratio. In direct gear, the rear-mounted gearbox NSG rotates as a block when the clutch KHR is open, and in overdrive the clutch KHR is closed. Due to the large transmission ratio, around 35% of the input torque is transmitted to the hub shell, with 65% remaining for the counter-torque to be absorbed by the main axle. This torque is to be switched by the KHR clutch when the NSG rear-mounted gearbox is switched from overdrive to direct gear. The axially stationary part of the KHR clutch is firmly connected to the ring gear HR5 via a side wall (see 3a) . The axial limitation of the movement space of the clutch KHR is ensured on the one hand by the center sleeve MH and the output-side flange of the main axle HA. The axially movable part of the clutch KHR is directly connected to the main axle HA via the splines and is controlled by a shift finger from the control groove NHR in the shift drum ST. To open the clutch KHR, both the static friction between the displaceable clutch part and the main axis HA and the static friction between the adjacent surfaces of the clutch teeth must be overcome.

In the following, exemplary embodiments of the invention are described as applied within an overall device such as a clutch assembly or a system.

The torque-dependent axial forces are preferably generated in the opening direction of the axial clutch by a twisted position of the normal of the effective surfaces of the clutch teeth engaging in one another. The normals advantageously have such an angle β _K with the orthogonal to the axis of rotation of the transmission that, in the limiting case, a compensation of the frictional forces against the axial forces is achieved, according to equation /10/. If, however, only a partial compensation is made by a lower angular position of the normal, the control groove NHR, as in 3b shown, maintained. In the case of complete compensation or overcompensation by the helical gearing(s), the control groove NHR is rotated over the entire angle of rotation of the shift drum, as in 3c shown, with a constant width adapted to the shift fingers, so that the axial clutch remains securely closed when closed.

Another preferred embodiment of the present invention is in 4d shown. An axial clutch that can be shifted under load relates to a block revolving clutch that connects or disconnects two rotating parts with one another in a switching manner. It is, for example, a K70 clutch of a ten-speed hub gear from the German patent application DE 10 2021 129 423 A1 . The reference symbols used there are also used and supplemented here.

The K70 clutch connects a sun gear SR3 with a center sleeve MH, which also serves as a dividing bridge for the output-side part of an input transmission EGG. The movable part of the axial coupling is guided inside the middle sleeve MH by means of a spline so that it can move axially and is secured against rotation. The axially fixed part of the K70 clutch is housed in one piece on the output side of the sun gear SR3, which is guided closely in the axial direction. The moving part of the K70 clutch has an annular collar on the inside, which is used as a contact surface for a sliding ring that is guided on the main axis HA so that it can move axially and is secured against rotation.

The sliding ring is designed in such a way that the collar inside the movable clutch part K70 is guided in a freely rotating manner by a non-rotating stop disk that is spring-loaded and an axially offset stop surface of the sliding ring. The axial length of the paragraph is only slightly greater than the wall thickness of the inner annular collar of the movable clutch part. This avoids a clamping effect. A spring F71 acting in the opening direction secures the open position of clutch K70. A spring F70 acting in the closing direction supports the closing process of the clutch K70, since a lane change takes place in a control groove N70 on a shift drum ST.

If the static friction is only partially compensated, the sliding ring together with the thrust washer AS is installed rotated by 180 degrees and the spring F71 is omitted. Because now, supported by the axial forces, clutch K70 is opened against the spring force of spring F70. This also applies in the same way to the other examples.

In the case of the K70 clutch, a slanted spline and the clutch spline can be used in combination. The clutch teeth are preferably positioned at an angle here.

A further preferred embodiment of the present invention of an axial clutch shiftable under load relates to a clutch which is seated on the outer surface of a rotating sleeve which is arranged coaxially with respect to a main axis HA and which is actuated through this sleeve. An example of this design is the clutch 90 from the German patent application DE10 2018 007 326 A1 . In contrast to the prior art, in the present case and advantageously, the axial clutch, which is assisted by the helix angle during decoupling, is controlled selectively via at least one actuator. The compensation of the frictional forces can preferably be divided between the splines and/or the clutch teeth.

Another preferred embodiment of the present invention of an axial clutch that can be shifted under load relates to an axial clutch that works as an anti-rotation lock for a sun gear and has to be shifted under load. Couplings of this type can be found in the German patent application DE 10 2018 007 326 A1 and in the German patent application DE 10 2021 129 423 A1 , in particular the local clutches 40 and K40 of the input transmission EGG. In the first sub-gear, this input gear has a complete planetary gear with sun, planet and ring gear, the drive of which is the planet carrier and the output takes place via the ring gear. To change gears, the sun gear is locked against rotation or released. A tracking lock is required for this. When releasing the axially closely guided sun gear, the axial clutch must be opened under load.

According to the invention, the movable clutch part is now guided inside the sun gear in this exemplary clutch assembly or this exemplary system or in this exemplary overall device concentrically, non-rotatably and axially displaceably by means of splines, and to support the shifting process, the splines are preferably guided with a helix angle β _S and the effective surfaces of the clutch teeth are rotated by the angle β _K relative to an imaginary plane through the axis of rotation 210 .

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 done as in 5a shown, made with screws that engage in holes matched to the screw. Shown in 5c is the control of the axial clutch for the case of compensation and overcompensation of the frictional forces. The activation of the axial clutch corresponds to that of clutch K70. In the event of undercompensation, what is described there also applies here.

The present invention comprises an axial clutch with a first shaft and with a second shaft, which can be coupled and uncoupled in a controlled manner by an actuator, comprising a first pair of active surfaces WP1 with a first helix angle β ₁ and a second pair of active surfaces WP2 with a second helix angle β ₂ , where the first helix angle β ₁ covers an angle range from 0 degrees to 30 degrees and the second helix angle β ₂ covers an angle range from 0 degrees to 30 degrees and the sum of the helix angles β ₁ plus β ₂ , via which a torque-dependent axial force AGES can be generated and which the Supports the decoupling process of the axial coupling under load, covering an angle range of 3 degrees to 35 degrees.

Axial clutches are preferably used in two-wheelers that are driven exclusively by a pedal or manual drive, or also in two-wheelers that are additionally or exclusively motor-driven, in particular in e-bikes or bicycles and cargo bikes supported by an electric motor.

character list

• 1a: eine Axialkupplung im Stand der Technik mit geradverzahnter Steckverzahnung und orthogonaler Kupplungsverzahnung,
• 1b: eine erfindungsgemäße Axialkupplung mit schrägverzahnter Steckverzahnung und gegenüber einer gedachten Ebene durch die Rotationsachse gekippten Wirkflächen der Kupplungszähne,
• 1c: eine orthogonale Sicht in Richtung Rotationsachse mit einer Ausschnittsmarkierung,
• 1d: den Ausschnitt aus 1c ergänzt durch die Skizze eines stumpfen Doppelkeils als Modell für das bewegliche Kupplungsteil, welches zur Herleitung der Auslegungsverfahren verwendet wird,
• 1e: ein Modell der Kupplungsbaugruppe aus 1b, ergänzt durch einen Kräfteplan bei einem Drehmoment, welches das Sonnenrad von links betrachtet im Uhrzeigersinn drehen möchte,
• 2: den schematischen Getriebeschnitt einer Hinterradnabe aus der DE 10 2021 129 423 A1 ,
• 3a: die Einbausituation einer Kupplung KHR einer bevorzugten Ausführungsform der Erfindung,
• 3b: eine bevorzugte Ausführungsform einer Steuernut NHR für den Fall der Unterkompensation,
• 3c: eine weitere bevorzugte Ausführungsform einer Steuernut NHR für den Fall der Kompensation bzw. Überkompensation der Axialkräfte,
• 3d: die Einbausituation einer Kupplung KSR, die indirekt mittels einer Kupplung KHR über schräggestellte Lauf- und/oder Steckverzahnung automatisch geschaltet wird,
• 4a: eine bevorzugte Ausführungsform einer unter Last schaltbaren Kupplung K70, die eine Blockumlaufkupplung ist,
• 4b: eine bevorzugte Ausführungsform einer Abwicklung der Steuernut N70 auf der Schalttrommel ST und die Raststellungen KR der Schaltstellungen G1-G10,
• 4c: eine bevorzugte Ausführungsform einer Steuernut N70 für den Fall der Unterkompensation,
• 5a: eine bevorzugte Ausführungsform einer Einbausituation einer Kupplung K40 in einer isolierten und detaillierteren Darstellung bei Kompensation bzw. Überkompensation der Axialkräfte,
• 5b: eine bevorzugte Ausführungsform einer Steuernut N40 für den Fall der Unterkompensation,
• 5c: eine bevorzugte Ausführungsform einer Steuernut N40 für den Fall der Kompensation bzw. Überkompensation,
• 6: eine bevorzugte Ausführungsform einer Einbausituation einer Kupplung K90 und
• 7: eine bevorzugte Ausführungsform eines Ausschnitts eines Eingangsgetriebes EGG mit schrägverzahnter Laufverzahnung.

Exemplary embodiments of the invention and the derivation of the underlying dimensioning method are described below with reference to figures. These figures show:

• 1a : a prior art axial clutch with spur spline and orthogonal clutch spline,
• 1b : an axial clutch according to the invention with helical splines and effective surfaces of the clutch teeth that are tilted relative to an imaginary plane through the axis of rotation,
• 1c : an orthogonal view towards the axis of rotation with a clip marker,
• 1d : the snippet off 1c supplemented by the sketch of a blunt double wedge as a model for the moving clutch part, which is used to derive the design process,
• 1e : a model of the clutch assembly 1b , supplemented by a force plan for a torque that wants to turn the sun gear clockwise when viewed from the left,
• 2 : the schematic gear section of a rear wheel hub from the DE 10 2021 129 423 A1 ,
• 3a : the installation situation of a clutch KHR of a preferred embodiment of the invention,
• 3b : a preferred embodiment of a control groove NHR for the case of undercompensation,
• 3c : another preferred embodiment of a control groove NHR for the case of compensation or overcompensation of the axial forces,
• 3d : the installation situation of a KSR clutch, which is automatically shifted indirectly by means of a KHR clutch via inclined running and/or spline teeth,
• 4a : a preferred embodiment of a clutch K70 that can be shifted under load, which is a block rotating clutch,
• 4b : a preferred embodiment of a development of the control groove N70 on the shift drum ST and the locking positions KR of the shift positions G1-G10,
• 4c : a preferred embodiment of a control groove N70 for the case of undercompensation,
• 5a : A preferred embodiment of an installation situation of a K40 clutch in an isolated and more detailed representation with compensation or overcompensation of the axial forces,
• 5b : a preferred embodiment of a control groove N40 for the case of undercompensation,
• 5c : a preferred embodiment of a control groove N40 for the case of compensation or overcompensation,
• 6 : a preferred embodiment of an installation situation of a clutch K90 and
• 7 : A preferred embodiment of a section of an input transmission EGG with helical gearing.

6. Detailed description

The same components have the same reference numbers.

1a shows a prior art axial clutch with straight splines and orthogonal clutch teeth.

1b shows an axial clutch according to the invention with helical splines and effective surfaces of the clutch teeth that are tilted relative to an imaginary plane through the axis of rotation.

A first shaft 200 applies a torque M1 to the axial clutch 100 . The first shaft 200 is connected to a spur gear 120 either via gearing (not shown) or fixed.

In the coupled state of the axial clutch 100, a second shaft 220 takes over a torque M2, the amount of which is equal to M1.

The first shaft 200 and the second shaft 220 are, in principle, supported in a gear arrangement so as to be rotatable about an axis of rotation 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 fixed axially and non-rotatably.

The axial clutch 100 comprises a clutch ring 130 which, in the engaged state, engages in a clutch toothing 114 which is connected to a plug-in toothing 113 in a torsionally fixed manner or alternatively in one piece. The spline 113 is axially displaceable and non-rotatably connected to the spur gear 120 . The spline 113 includes at least one tooth 112, which has a helix angle β _S of 0 degrees to 30 degrees.

The coupling ring 130 comprises at least one asymmetrical tooth 131 which, in the engaged state, engages in 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 1d ) on. The ramp 133 has a helix angle β3 in the range from 45 degrees to 75 degrees with respect to an imaginary plane through the axis of rotation 210 (see FIG 1d ). This serves to safely overcompensate for the frictional forces so that the clutch can work as a freewheel. The Flank 132 has a helix angle β _K in the range from 0 degrees to a maximum of 30 degrees with respect to an imaginary plane through the axis of rotation 210 (see 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 _S each total 3 degrees to 35 degrees, in particular 5 degrees to 22 degrees and preferably 7 degrees to 20 degrees provide the axial forces A _S and A _K required to support a separation of the axial coupling 100 under load (see 1e) . The axial forces A _S and A _K support the opening of the axial clutch 100 in the opening direction x as soon as the clutch ring 130 is actuated by an actuator 140 (see 1e) is controlled in the opening direction x.

The sum of the respective 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 .

The axial force components A _S and A _K help to overcome the frictional forces R _S and R _K . In this way, undercompensation, compensation or overcompensation of the frictional forces R _S and R _K can be generated. The axial forces A _S and A _K are dependent on the difference between the magnitude of the input torque acting on the axial clutch 100 of the gear stage to be disengaged and the magnitude of the output torque of the gear stage to be disengaged.

The axial forces A _S and A _K support the opening of the axial clutch 100 in the opening direction x as soon as it is actuated by an actuator 140 (see 1e) is controlled in the opening direction x.

Steel and aluminum or steel and steel are preferably used as materials. 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 fibers, can also be considered, if possible to be able to realize small friction angles ρ _S and ρ _K , in which safe coupling and decoupling between the movable coupling part 110 and the coupling ring 130 is to be ensured even under high loads and severe external shocks.

The axial clutch comprises a material pairing between the clutch ring 130 and the clutch teeth 114 and between the spline 113 and the guide of the spur gear 120 a material pairing made of steel, aluminum, aluminum multi-component bronze, bronze, beryllium copper, brass, industrial ceramics or solid or surface-coated components carbon or carbon fibers, to to ensure safe engagement and disengagement between the lock and the clutch ring.

On the movable coupling part 110 is optional springs 111; 111' exerts a force closing the clutch in relation to a fixed housing part or in relation to an axially rigid component (see 1e) .

The movable coupling part 110 is guided in an axially movable manner via the spur gear 120 . The spur gear 120 has internal splines with a helix angle β _S .

1c shows an orthogonal view in the direction of the axis of rotation with a clip marker.

1d shows the section 1c supplemented by the sketch of a blunt double wedge as a model for the moving clutch part, which is used to derive the design process.

On the movable coupling part 110 with splines, the helix angle β _S is shown using the example of a tooth 112 of the spline and the helix angle β _K is shown using the example of a tooth of the coupling teeth 114, which together form a derived axially movable wedge model, which is shown in 1e is shown in more detail.

1e shows a model of the clutch assembly 1b , supplemented by a force plan for a torque that wants to turn the sun gear clockwise when viewed from the left. There you will find the double wedge, which represents a problem analogous to an axial clutch with a spline and a clutch spline. In this analogy, the wedge is the moving part of the coupling. This has two helical active surfaces, for which the mathematical derivation is analogous (i = S for the splines and i = K for the clutch teeth). In the following, all forces are considered in absolute terms. The normal forces N _i act on the effective surfaces. This results in the frictional forces that act parallel to the tooth flanks: $${\text{R}}_{\text{i}}={\mathrm{\mu }}_{\text{i}}*{\text{N}}_{\text{i}}$$

The tangential component of the normal force, N _i * cos(βi), must be applied by the applied tangential force F _i . In addition, the frictional force also has a component R _i * sin(βi) in the direction of the applied tangential force, which must also be applied. So the following applies: $${\text{f}}_{\text{i}}={\text{N}}_{\text{i}}*\text{cos}\left({\mathrm{\beta }}_{\text{i}}\right)+{\text{R}}_{\text{i}}*\text{sin}\left({\mathrm{\beta }}_{\text{i}}\right)$$

Using the equations /1/ and /2/, the following relation can be derived: $${\text{N}}_{\text{i}}={\text{f}}_{\text{i}}/\left(\text{cos}\left({\mathrm{\beta }}_{\text{i}}\right)+{\mathrm{\mu }}_{\text{i}}*\text{sin}\left({\mathrm{\beta }}_{\text{i}}\right)\right)$$

The applied tangential force F _i depends on the applied torque and the effective pitch circle radius of the gearing: $${\text{f}}_{\text{i}}={\text{m/r}}_{\text{i}}$$

On the one hand, a component of the normal force, A _i = N _i * sin(βi), and a component of the frictional force, R _i * cos(βi), act in the opposite direction in the axial direction. The resulting axial force X _i is thus: $${\text{X}}_{\text{i}}={\text{N}}_{\text{i}}*\text{sin}\left({\mathrm{\beta }}_{\text{i}}\right)-{\text{R}}_{\text{i}}/\text{cos}\left({\mathrm{\beta }}_{\text{i}}\right)$$

bestimmen. Für den Spezialfall A_i = 0 ergibt sich dann: $${\mathrm{\beta }}_{\text{i}}=\text{arctan}\left({\mathrm{\mu }}_{\text{i}}\right)$$

The axial force X _i can be calculated using equations /1/ to /5/ $${\text{X}}_{\text{i}}=\text{M}/{\text{right}}_{\text{i}}*\left[\text{sin}\left({\mathrm{\beta }}_{\text{i}}\right)-{\mathrm{\mu }}_{\text{i}}*\text{cos}\left({\mathrm{\beta }}_{\text{i}}\right)\right]/\left[\text{cos}\left({\mathrm{\beta }}_{\text{i}}\right)+{\mathrm{\mu }}_{\text{i}}*\text{sin}\left({\mathrm{\beta }}_{\text{i}}\right)\right]$$

determine. For the special case A _i = 0 this results in: $${\mathrm{\beta }}_{\text{i}}=\text{arctan}\left({\mathrm{\mu }}_{\text{i}}\right)$$

A helix angle that satisfies this relation is also called the friction angle ρ _i .

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.: $${\text{X}}_{\text{TOTAL}}={\text{X}}_{\text{S}}+{\text{X}}_{\text{K}}$$

According to the invention, in the case of complete compensation of the frictional forces, the sum of the axial forces should be equal to zero, i.e. X _GES =0. It follows from this: $$\begin{array}{l}\left[\text{sin}\left({\mathrm{\beta }}_{\text{K}}\right)-{\mathrm{\mu }}_{\text{K}}*\text{cos}\left({\mathrm{\beta }}_{\text{K}}\right)\right]/\left[\text{cos}\left({\mathrm{\beta }}_{\text{K}}\right)+{\mathrm{\mu }}_{\text{K}}*\text{sin}\left({\mathrm{\beta }}_{\text{K}}\right)\right]\\ =-{\text{right}}_{\text{K}}/{\text{right}}_{\text{S}}*\left[\text{sin}\left({\mathrm{\beta }}_{\text{S}}\right)-{\mathrm{\mu }}_{\text{S}}*\text{cos}\left({\mathrm{\beta }}_{\text{S}}\right)\right]/\left[\text{cos}\left({\mathrm{\beta }}_{\text{S}}\right)+{\mathrm{\mu }}_{\text{S}}*\text{sin}\left({\mathrm{\beta }}_{\text{S}}\right)\right]\end{array}$$

mit: $$\text{a}={\mathrm{\mu }}_{\text{K}}$$

$$\text{b}=-{\text{r}}_{\text{K}}/{\text{r}}_{\text{S}}*\left[\text{sin}\left({\mathrm{\beta }}_{\text{S}}\right)-{\mathrm{\mu }}_{\text{S}}*\text{cos}\left({\mathrm{\beta }}_{\text{S}}\right)\right]/\left[\text{cos}\left({\mathrm{\beta }}_{\text{S}}\right)+{\mathrm{\mu }}_{\text{S}}*\text{sin}\left({\mathrm{\beta }}_{\text{S}}\right)\right]$$

β _K is thus dependent on β _S , where a solution can be found by: $${\mathrm{\beta }}_{\text{K}}=2*\text{arctan}\left(\left[\text{a}*\text{b}+{\left(\left({\text{a}}^2+1\right)*\left({\text{b}}^2+1\right)\right)}^{1/2}-1\right]/\left[\text{a}+\text{b}\right]\right)$$

With: $$\text{a}={\mathrm{\mu }}_{\text{K}}$$

$$\text{b}=-{\text{right}}_{\text{K}}/{\text{right}}_{\text{S}}*\left[\text{sin}\left({\mathrm{\beta }}_{\text{S}}\right)-{\mathrm{\mu }}_{\text{S}}*\text{cos}\left({\mathrm{\beta }}_{\text{S}}\right)\right]/\left[\text{cos}\left({\mathrm{\beta }}_{\text{S}}\right)+{\mathrm{\mu }}_{\text{S}}*\text{sin}\left({\mathrm{\beta }}_{\text{S}}\right)\right]$$

A sufficiently good approximation could be determined empirically for the usual helix angles, given by: $${\mathrm{\beta }}_{\text{K}}={\mathrm{\rho }}_{\text{K}}-\left({\text{right}}_{\text{K}}/{\text{right}}_{\text{S}}\right)*\left({\mathrm{\beta }}_{\text{S}}-{\mathrm{\rho }}_{\text{S}}\right)$$

The derivation is analogous for the other embodiments according to the invention, including an inclined position of a plug-in toothing and a running toothing or a running toothing and a clutch toothing.

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 system of a transmission.

2 shows a rear wheel hub with a multi-speed transmission, which is formed from planetary gears, which engage in helical grooves N24 to N74, NSR, NHR in a shift drum by means of axially displaceable sliding rings with shift fingers, the shift fingers also being in radial slots oriented parallel to the transmission axis of a enclosing slot are guided on the hollow main axis HA and as a result different gears can be shifted, in that radially extending clutch rings, which are fitted with axial teeth according to the invention, are to be released according to a respective gear step by means of the sliding rings and are spring-loaded, with the multi-speed gearbox arranged on the main axis starting from a drive sleeve AH and starting from a five -Gear transmission as an input transmission EGG, which consists of a two-stage and a subsequent single-stage planetary gear, as well as a subsequent single-stage planetary gear that converts it into high speed as a rear-mounted gear NSG, which includes a It is arranged on a center sleeve MH and, when coupled, connects the first five gears to a hub sleeve NH or, with a fixed ring gear HR4, transmits these gears in a doubled coupled manner to the hub sleeve NH, with the 10 gear stages formed in this way being transmitted with 7 clutches K20, K30, ... , KHR are connected in such a way that only a few of the planetary gearsets are engaged in a rolling manner under load.

At least two of the planetary gears PR1, PR2 are identical to one another and these have annular ring gears HR2-3 connected to one another, and the planet carriers PT2, PT3 of these two planetary gears are supported against one another axially and radially with an intermediate bearing ZL.

The sun gear SR5 of the rear-mounted gearbox NSG is mounted on the middle sleeve MH and is axially displaceable and non-rotatably connected to the hub sleeve NH, and can also be coupled according to the invention with a sun gear clutch KSR to the output-side planetary carrier PT3 of the input gear EGG, and the ring gear HR4 of the rear-mounted gearbox NSG is close to the main axis HA can be coupled according to the invention with a ring gear clutch KHR.

3a shows the installation situation of a clutch KHR of a preferred embodiment of the invention.

3b shows a preferred embodiment of a development of the control groove NHR on the shift drum ST and the detent positions KR of the shift positions G1-G10.

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

In 3d Another simplified embodiment of the rear-mounted gearbox is shown, in which the sun gear clutch KSR is not shifted from the shift drum ST, but rather the thrust forces of a helical gearing of the rear-mounted gearbox and/or the helical gearing of the splined gearing of the torque transmitter in the hub shell NH are used in such a way that the Sun wheel clutch KSR is always automatically opened or closed against spring force, opposite to the actively switched 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 gearbox therefore rotates as a block. In the preferred variant, the helical running gear and/or the helical spline is oriented in such a way that the sun gear SR5 moves away from the stationary part of the clutch KSR under load due to the axial force and the ring gear HR5 experiences an opposing force, which acts against the hub shell with the axial support bearing ASL NH is supported. The torque transmitter D5 is non-rotatably but axially displaceable with the hub shell NH and non-rotatably and axially firmly connected to the sun gear SR5. When the rear-mounted gearbox NSG is stationary or in direct gear during operation, the sun gear clutch KSR is by means of the return spring F5, which presses the torque transmitter D5 axially against the hub sleeve NH supported, closed. This movement is limited by a stop in the NH hub shell. The axial forces of the ring gear HR5 are absorbed directly via an axially loadable roller bearing between the hub sleeve NH and the ring gear HR5 and/or via an axial support element ASE on the main axis HA. The shifting process is supported by the fact that when the ring gear clutch KHR is closed, the sun gear SR5 overtakes the planet carrier PT5 and thus switches to freewheeling mode, with the orientation of the clutch teeth also ensuring that the clutch KSR is opened. The drive of the hub sleeve NH is changed from the middle sleeve MH to the sun gear SR5 without any load interruption.

4a shows a preferred embodiment of a clutch K70 according to the invention that can be shifted under load. It is a block rotating clutch that connects or disconnects two rotating parts in a switching manner. It is, for example, the K70 clutch of a 10-speed hub gear from the patent application DE 10 2021 129 423.1 . This coupling is designed here according to the invention. The reference symbols 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 middle sleeve MH, which also serves as a partial bridge for the output-side part of the input transmission. The moving part of the coupling is guided inside the middle sleeve by means of a spline so that it can move axially and is secured against rotation. The axially fixed part of the clutch is housed in one piece on the output side of the axially narrow guided sun gear SR3. The moving part of the coupling has an annular collar on the inside, which is used as a contact surface for the sliding ring, which is guided on the main axis so that it can move axially and is secured against rotation. The sliding ring is designed in such a way that the collar is guided freely rotating inside the movable coupling part by a non-rotating stop ring that is spring-loaded and an axially offset stop face of the sliding ring. The axial length of the paragraph is only slightly greater than the wall thickness of the inner ring-shaped collar of the movable clutch part. This avoids a clamping effect. The spring F71 acting in the opening direction secures the open position of the clutch. The spring F70 acting in the closing direction supports the closing process of the clutch, since a lane change takes place in the control groove N70 on the shift drum ST. If the static friction is only partially compensated, the sliding ring together with the thrust washer is installed rotated by 180 degrees and the spring F71 is omitted. Because now, supported by the axial forces, the K70 clutch is opened against the spring force of the F70. This also applies in the same way to the other examples.

With the K70, a slanted spline and the clutch spline can be used in combination. The clutch teeth are preferably positioned at an angle here.

The axial forces A _S and A _K (see respectively 1e) support the opening of the axial clutch 100 in the opening direction x (also in 1e shown) as soon as this is controlled by an actuator 140.

4b shows a preferred embodiment of a development of the control groove N70 on the shift drum ST and the detent positions KR of the shift positions G1-G10 for the case of under-compensation.

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

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 drive is a planetary carrier and the output is via a ring gear. To change gears, a sun gear SR2 is locked in rotation or released. A tracking lock is required for this. When releasing sun wheel SR2, clutch K40 must be opened under load. Within the sun gear SR2, the movable clutch part is guided concentrically to this, by means of a spline SST in a rotationally fixed, axially displaceable manner. To support the shifting process, the splines SST are rotated by the angle β _K with a helix angle β _S and/or the effective surfaces of the clutch teeth relative to an imaginary plane through the axis of rotation of the transmission. The non-rotatable and axially fixed part of the K40 clutch is connected to the main axle HA via a spline SST. The axial locking is preferred, as in 5a shown, made by means of screws, which engages in holes adapted to the screw. Shown in 5c is the control of the axial clutch for the case of compensation and overcompensation of the frictional forces. The control of clutch K40 corresponds to that of clutch K70. In the event of undercompensation, what is described there also applies here.

The axial forces A _S and A _K (see respectively 1e) support the opening of the axial clutch 100 in the opening direction x (also in 1e shown) as soon as this is controlled by an actuator 140.

5b shows a preferred embodiment of a groove N40 for the case of undercompensation.

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

6 shows a preferred embodiment of an installation situation of a clutch K90.

A preferred embodiment of the present invention for such a transmission is in 6 shown schematically. The output-side clutch K90 of the rear-mounted gearbox NSG shown there is controlled via two cascaded shift fingers, one of which is guided in a groove N94 in the shift drum ST with low play and the second of which passes through the rotating output sleeve H and a clutch ring 130 controls. Two springs F90 and F91 center the stacked shift fingers. When the K90 clutch is closed, 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 circlip SR fixes its axial play on the output sleeve H.

According to another preferred embodiment, an axial coupling 100 is placed closer to a main axis. The axial clutch 100 lies in a rotating sleeve H. Helical gearing used to lighten the switch is arranged between a planetary carrier PT and an output-side clutch half. Shift fingers SF and centering springs also act on this. Stacking of shift fingers SF is therefore not necessary, which considerably simplifies the design and manufacture.

At the clutches in the transmission according to the DE 10 2018 007 A1 the shifting forces of some K50, K90 clutches are fed to the shifting rings in several stages in the output-side transmission block, with shifting elements passing through sleeves that are moving along, so that axial forces that are difficult to control occur on the shifting elements if the gear toothing is inclined or are necessary for shifting. There are therefore only slight inclinations of a few degrees provided there.

The axial forces A _S and A _K (see respectively 1e) support the opening of the axial clutch 100 in the opening direction x (also in 1e shown) as soon as this is controlled by an actuator 140.

7 shows a section of the input transmission EGG with at least one axial clutch 100 according to the invention. The two symmetrical planetary transmissions with the planetary gears PR2 and PR3 are provided with helical gearings 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 clutches K40, K50 according to the invention. The reaction forces acting on the interconnected ring gears HR2 and HR3 via the planetary gears PR2, PR3 compensate each other and keep them centered on one another as a so-called flying ring pair. Closing springs F40, F50 oppose the opening forces A _S and A _K . As a result, relatively moderate shifting forces are required on the shift fingers, which each have a loose fit in control grooves N40, N50, in order to switch over the axial clutches even at high transmission loads.

The release springs F41, F51 are dimensioned in such a way that the sliding rings secure the open clutch positions, even if no auxiliary force is available from the helical gearing when there is no or little gear load.

It is advantageous for the sun gears and the associated clutch ring to be manufactured in one piece. In the designated rear wheel transmission, a clutch travel of only about 1.5 mm between the closed and open states is sufficient.

at 7 the axial forces of the running teeth and those of the clutch teeth are used. Only clutch K40 is opened under load. Clutch K50 goes into freewheeling mode when clutch K70, the block rotating clutch, is closed. Sun gear SR3 then becomes simple carried along in the direction of rotation. For this reason, the bevel on the K50 clutch could be dispensed with. The control then remains as it is in the 2 you can see. This provides an embodiment that makes it possible to find good solutions when space is difficult. In this case, the axial forces of the K50 clutch are undercompensated.

The K40 clutch works as a rotation lock and the K50 clutch works as a backstop.

$${\mathrm{\beta }}_{\text{K}}={\mathrm{\rho }}_{\text{K}}=\mathrm{6,84}\text{ Grad}\text{, }{\mathrm{\beta }}_{\text{L}}={\mathrm{\rho }}_{\text{L}}=\mathrm{0,86}\text{ Grad},$$

$${\mathrm{\beta }}_{\text{K}}=3\text{ Grad}=>{\mathrm{\beta }}_{\text{L}}=\mathrm{6,44}\text{ Grad sowie}$$

$${\mathrm{\beta }}_{\text{K}}=5\text{ Grad}=>{\mathrm{\beta }}_{\text{L}}=\mathrm{3,54}\text{ Grad}.$$

In a preferred embodiment, the helical gearing is located on the one hand on the running gears and on the other hand also in the clutch gearing. This is due to the geometric and tribological situation, since only the K40 clutch has to be switched under load. With an assumed rolling friction coefficient of the running 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, an average pitch circle radius r _K of 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 the frictional forces are compensated according to the inventive method: $${\mathrm{\beta }}_{\text{K}}=0=>{\mathrm{\beta }}_{\text{L}}=10.9\text{Degree}$$

$${\mathrm{\beta }}_{\text{K}}={\mathrm{\rho }}_{\text{K}}=6.84\text{Degree}\text{,}{\mathrm{\beta }}_{\text{L}}={\mathrm{\rho }}_{\text{L}}=0.86\text{Degree},$$

$${\mathrm{\beta }}_{\text{K}}=3\text{Degree}=>{\mathrm{\beta }}_{\text{L}}=6.44\text{degree as well}$$

$${\mathrm{\beta }}_{\text{K}}=5\text{Degree}=>{\mathrm{\beta }}_{\text{L}}=3.54\text{Degree}.$$

It can be seen that the sum of the friction angles ρ _K and ρ _L only corresponds to the angle sum that is actually required for individually compensated gearing.

The axial force AGES supports the opening of the axial clutch 100 in the opening direction x as soon as it is actuated by an actuator 140 (see 1e) is controlled in the opening direction x.

the 2 until 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 is subjected to a closing force by a closing spring 111, 111' and also has 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 side wall of the groove exerts a shifting force against the closing force on the shift finger and thereby opens the axial clutch 100 and thus produces a different shift state of the respective gear stage, characterized that the gearing of the planetary gear is designed as a helical gearing so that torque-dependent axial forces AGES occurring under a load in the opening direction of the axial clutch 100, comprising the clutches K40, K50, and which largely compensate for the static friction occurring between the contacting clutch teeth and acting in the opposite direction to a separation of the axial clutch 100, with the resultant reaction forces occurring on the clutch elements being absorbed by support elements on adjacent transmission components.

100

Axialkupplung

110

beweglicher Kupplungsteil

111; 111'

Feder

112

wenigstens ein Zahn der Steckverzahnung

113

Steckverzahnung

114

Kupplungsverzahnung

120

Stirnrad

130

Kupplungsring

131

asymmetrische Zähne

132

steile Flanke der Kupplungsverzahnung

133

flache Rampe der Kupplungsverzahnung

140

Aktuator

200

erste Welle

220

zweite Welle

210

Rotationsachse

300

Schiebering mit Bund

600

Laufverzahnung

WP1

erste Wirkflächenpaarung

WP2

zweite Wirkflächenpaarung

MP1

erste Materialpaarung

MP2

zweite Materialpaarung

ps

Reibungswinkel Steckverzahnung

ρK

Reibungswinkel Kupplungsverzahnung

ρL

Reibungswinkel Laufverzahnung

βS

Schrägungswinkel von 113

βK

Schrägungswinkel von 114

βL

Schrägungswinkel von 600

β1

erster Schrägungswinkel

β2

zweiter Schrägungswinkel

β3

Schrägungswinkel von 133

µS

Reibungswert Steckverzahnung

µK

Reibungswert Kupplungsverzahnung

µL

Reibungswert Laufverzahnung

rS

Teilkreisradius Steckverzahnung

rK

mittlerer Teilkreisradius Kupplungsverzahnung

rL

Teilkreisradius Laufverzahnung

NS, NK

Normalkraft Steck-/Kupplungsverzahnung

RS, RK

Reibkraft Steck-/Kupplungsverzahnung

FS, FK

Tangentialkraft Steck-/Kupplungsverzahnung

AS, AK

axiale Komponente von N_S, N_K

AGES

Summe der durch Schrägstellung von Zahnrädern Erzeugten Axialkräfte

XS, XK

Effektive Axialkraft an Steck-/Kupplungsverzahnung

XGES

Summe der effektiven Axialkräfte

x,

Öffnungsrichtung

M1, M2

Drehmoment

M

Betrag des Drehmoments M1/M2

AH

Antriebshülse

HA

Hauptachse

MH

Mittelhülse

NH

Nabenhülse

ST

Schalttrommel

N20, N30, N40, N50, N60, N70, N80, N90, NSR, NHR

Steuernut(en)

L

Lager von MH

PT1-2, PT3/PT3-4, PT5

Planetenträger

ZL

Zwischenlager PT1-2 zu PT3-4

SR1 - SR5

Sonnenräder

PR1 - PR5

Planetenräder

HR2-3, HR5

Hohlräder

F5

Rückstellfeder

F40, F50

Schließfeder

F41, F51

Lösefeder

F70, F71

Feder

ASL

axiales Stützlager

ASE

axiales Stützelement

D5

Drehmomentübertrager

KR

Raststellungen

AS

Anlaufscheibe

IGH

Integrated Gear Hub

HR

Hohlrad

PR

Planetenrad

SR

Sonnenrad

EGG

Eingangsgetriebe

NSG

Nachschaltgetriebe

G1 ... G10

Schaltstellungen

K20, K30, K40

Kupplungen erstes Teilgetriebe

K50, (K60), K70, K80

Kupplungen zweites Teilgetriebe

KSR

Sonnenradkupplung im NSG

KHR

Hohlradkupplung im NSG

SST

schrägverzahnte Steckverzahnung

SF

Schaltfinger

K

Kulisse

H

Hülse

The list of reference numbers refers to the accompanying figures.

100

axial coupling

110

moving part of the clutch

111; 111'

Feather

112

at least one tooth of the spline

113

spline

114

clutch teeth

120

spur gear

130

clutch ring

131

asymmetrical teeth

132

steep flank of the clutch teeth

133

flat ramp of clutch teeth

140

actuator

200

first wave

220

second wave

210

axis of rotation

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 spline

ρK

Friction angle of clutch teeth

ρL

Friction angle of gear teeth

βS

helix angle of 113

βK

helix angle of 114

βL

Helix angle of 600

β1

first helix angle

β2

second helix angle

β3

helix angle of 133

µS

Friction value spline

µK

Coupling gear friction coefficient

µL

Friction value of gear teeth

rS

Pitch circle radius spline

rK

medium pitch circle radius clutch teeth

rL

Pitch circle radius running gear

NS, NK

Normal force plug-in/clutch teeth

RS, RK

Friction force plug-in/clutch gearing

FS, FK

Tangential force plug-in/clutch teeth

AS, AK

axial component of N _S , N _K

AGES

Sum of the axial forces generated by the skewing of gears

XS, XK

Effective axial force on plug-in/clutch splines

XGES

Sum of the effective axial forces

x,

opening direction

M1, M2

torque

M

Amount of torque M1/M2

AH

drive sleeve

HA

main axis

MH

center sleeve

NH

hub shell

ST

shift drum

N20, N30, N40, N50, N60, N70, N80, N90, NSR, NHR

control groove(s)

L

MH warehouse

PT1-2, PT3/PT3-4, PT5

planet carrier

ZL

Intermediate storage PT1-2 to PT3-4

SR1 - SR5

sun gears

PR1 - PR5

planet gears

HR2-3, HR5

ring gears

F5

return spring

F40, F50

closing spring

F41, F51

release spring

F70, F71

Feather

ASL

axial support bearing

ASE

axial support element

D5

torque transmitter

KR

detent positions

AS

thrust washer

ICJ

Integrated gear hub

MR

ring gear

PR

planet wheel

SR

sun gear

EGG

input gear

NSG

secondary gearbox

G1...G10

switching positions

K20, K30, K40

Couplings first partial transmission

K50, (K60), K70, K80

Couplings second partial transmission

KSR

Sun gear clutch in the NSG

KHR

Ring gear clutch in the NSG

SST

helical splines

SF

shift finger

K

backdrop

H

sleeve

Claims (15)

Axial clutch (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 pair of active surfaces (WP1) with a first helix angle (β ₁ ) and a second effective surface pairing (WP2) with a second helix angle (β ₂ ), the first helix angle (β ₁ ) an angular range of 0 degrees to 30 degrees relative to an imaginary plane through a rotation axis (210) and the second helix angle (β ₂ ) an angular range from 0 degrees to 30 degrees relative to an imaginary plane through an axis of rotation (210) and the sum of the helix angles (β ₁ ) plus (β ₂ ) via which a torque-dependent axial force (A _GES ) can be generated and which the uncoupling process of the axial clutch ( 100) supported under load, covers an angle range of 3 degrees to 35 degrees. Axial coupling (100) after claim 1 wherein the first pair of active surfaces (WP1) comprises a spline (113) with at least one tooth (112) and the first helix angle (β ₁ ) comprises a helix angle (β _S ) and the second pair of active surfaces (WP2) comprises a coupling tooth system (114) and the second helix angle (β ₂ ) comprises a helix angle (β _K ). Axial coupling (100) after claim 2 , wherein a coupling ring (130) of the coupling toothing (114) has an asymmetrical toothing (131) with at least one tooth (112), comprising a flank (132) having the angle (β _K ). Axial coupling (100) after claim 3 , wherein the coupling ring (130) is designed to be controllable by an actuator (140) in an opening direction x, the axial force (A _GES ) supporting the opening of the axial coupling (100) in the opening direction x. Axial coupling (100) after claim 1 wherein the first pair of active surfaces (WP1) comprises a spline (113) with at least one tooth (112) and the first helix angle (β ₁ ) comprises a helix angle (β _S ) and the second pair of active surfaces (WP2) comprises a running gear (600) and the second helix angle (β ₂ ) comprises a helix angle (β _L ). Axial coupling (100) after claim 1 wherein the first pair of active surfaces (WP1) comprises a clutch tooth system (114) with at least one tooth (112), and the first helix angle (β ₁ ) comprises a helix angle (β _K ) and the second pair of active surfaces (WP2) comprises a running tooth system (600) and the second helix angle (β ₂ ) comprises a helix angle (β _L ). Axial clutch (100) according to one of the preceding claims, wherein the first pair of materials (MP1) of the first pair of active surfaces (WP1) and the second pair of materials (MP2) of the second pair of active surfaces (WP2) one of the material pairs of steel, bronze, beryllium copper, aluminum, aluminum multi-component bronze, Ceramic, carbon or carbon fibers included individually or in combination to ensure safe decoupling. 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 acted upon on the one hand by a closing spring (111, 111') with a closing force 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, its groove side wall exerts a shifting force against the closing force on the shift finger, thereby opening the axial clutch (100) and thus changing the shifting state of the produce the respective gear stage, characterized in that the gearing of the planetary gearing is designed as helical gearing in such a way that torque-dependent axial forces (A _GES ) occurring under a load are thereby transmitted in the opening direction of the axial clutch (100), comprising the clutches (K40 , K50), and which are largely compensated by the static friction occurring between the contacting clutch teeth and acting in the opposite direction to a separation of the axial clutch (100), with the resultant reaction forces occurring on the clutch elements being absorbed by support elements on adjacent transmission components. Clutch assembly comprising a planetary gear according to claim 8 , characterized in that sun gears (SR2, SR3) of two mirror-inverted, adjacent helical planetary gears which are connected to or released from a main axle (HA) by clutches (K40, K50), the reaction forces occurring in each case on the sun gears, which are the associated Planetary gears (PR2, PR3) are transmitted and forwarded by them to the associated ring gears (HR2-3), which are intercepted by the combination of each other. Clutch assembly comprising a planetary gear according to claim 8 , characterized in that the clutch rings are each connected to the clutch (K40, K50). Clutch assembly, comprising a planetary gear according to one of Claims 8 until 10 , characterized in that the helical gearing is a running gear whose helix angle is between 5 degrees and 12 degrees. Clutch assembly comprising a planetary gear according to claim 8 , characterized in that the helical gearing is a spline and the ring gear is supported with a support bearing on the hub sleeve and above it on the gear housing. Clutch assembly comprising a planetary gear according to claim 8 , characterized in that the shift finger, which is guided in the groove (N60/N80) with low tolerances in the control drum (ST), on the main axle (HA) on both sides axially between two springs, which are fixed on the main axle at the axial ends with support washers , is kept centered, drives a sun gear with helical teeth towards the planet carrier (PT) in an axially displaceable manner, which carries a clutch ring whose pedant is fixed axially to the rotatable sleeve (H) and forms the clutch (K60, K80) to be switched. Method for dimensioning the axial clutch according to one of the preceding claims in the following steps: • Determination of a coefficient of friction (µ ₁ ) between the active surface pair (WP1) involved in a first pair of materials (MP1) and a coefficient of friction (µ ₂ ) between the pair of active surfaces in a second pair of materials ( MP2) participating effective surface pair (WP2) by means of conventional methods on the object and / or taken from a table with material properties. • Determination of an effective pitch circle radius (r ₁ ) of the first effective surface pairing (WP1) and an effective pitch circle radius (r ₂ ) of the second effective surface pairing (WP2) from the design data. • Determination or specification of a first helix angle (β ₁ ) of the first pair of active surfaces (WP1) and a second helix angle (β ₂ ) of the second pair of active surfaces (WP2) from such design data that are load-independent, namely the coefficient of friction of the first pair of materials (MP1) and the second pair of materials (MP2) (μ ₁ ) and (μ ₂ ), the pitch circle radii of the first pair of effective surfaces (WP1) and the second pair of effective surfaces (WP2) (r ₁ ) and (r ₂ ), where z. B. in the case of complete compensation of the frictional forces, the angle (β ₂ ) can be determined as a function of (β ₁ ): $${\mathrm{\beta }}_2=2*\text{arctan}\left(\left[\text{a*b+}{\left(\left({\text{a}}^2+1\right)*\left({\text{b}}^2+1\right)\right)}^{1/2}-1\right]/\left[\text{a}+\text{b}\right]\right)$$

With: $$\text{a}={\mathrm{\mu }}_2$$

$$\text{b}=-{\text{right}}_2/{\text{right}}_1*\left[\text{sin}\left({\mathrm{\beta }}_1\right)-{\mathrm{\mu }}_1*\text{cos}\left({\mathrm{\beta }}_1\right)\right]/\left[\text{cos}\left({\mathrm{\beta }}_1\right)+{\mathrm{\mu }}_1*\text{sin}\left({\mathrm{\beta }}_1\right)\right].$$

Procedure for dimensioning the axial coupling according to Claim 14 , having an approximation for angles up to 30° with: $${\mathrm{\beta }}_2={\mathrm{\rho }}_2=\left({\text{right}}_2/{\text{right}}_1\right)*\left({\mathrm{\beta }}_1-{\mathrm{\rho }}_1\right)$$

and with: $${\mathrm{\rho }}_{\text{i}}=\text{arctan}\left({\mathrm{\mu }}_{\text{i}}\right).$$

Images