DE102022107914A1 - Pawl clutch of a switching device and method for dimensioning - Google Patents

Abstract

The present invention relates to a pawl clutch of a switching device, in which a pawl (1) with a locking lever (3) and a control lever (4) is pivotally mounted on a pivot pin (2) on a drive axle (6) and in each case when the control lever ( 4) is pivoted into a cam valley (NT) of a cam body (5), an arm of the locking lever (3) with an end face (SF) engages in a load-dependent, force-fitting manner in an internal toothing (ZI) of a gear wheel on a coupling surface (KF) and in each case, When the control lever (4) comes into contact with a cam (N) by adjusting the cam body (5), the locking lever (3) releases against a load-dependent friction force FK from the internal toothing (ZI). The end face (SF) and the dome surface (KF) are inclined at an angle (β) to a perpendicular to a vector emanating radially from the pawl pivot axis (8) and leading to the center of the end face (SF). The present invention further relates to a method for dimensioning a pawl clutch.

Description

1. Technical area

The invention relates to a pawl clutch of a switching device, in particular a planetary gear or a spur gear in a pedal-driven bicycle hub or in the hub of a motor-driven e-bike.

2. State of the art

From the EP 0 915 800 B1 A bicycle hub with several pawl clutches of associated planetary gears is known. On a hollow drive axle, two-armed pawls are mounted on pivot pins, of which one locking lever arm engages at the front in an asymmetrical internal toothing of a sun gear, supporting a drive torque, when the other lever arm is pressed by a spring as a control lever into a cam valley of a cam body is arranged inside the drive axle and can be moved into different gear positions using external switching means. The front side of the locking lever is oriented perpendicular to the direction of the locking force, which runs through the pivot center of the pivot pin. When unlocking shifting under load, considerable frictional forces on the front side as well as on the pivot pin have to be overcome in addition to a spring force, which often means that the shifting is delayed when a next gear has already been shifted, which in special circumstances leads to a complete blockage of the entire gear Gear arrangement leads. This problem occurs with pedal crank drives and is particularly serious when not only a pedal crank drive, but also a motorized additional drive is active during shifting.

Furthermore, from the DE 10 2013 112 788 B4 It is known that such a switching pawl arrangement is provided with a multi-stepped cam profile when switching under load due to the high frictional forces of the pawls and the cam bodies, so that the separating torque of the pawl with the angle of rotation of the cam over approximately 90 ° from a high initial value to a tenth of it falls off.

The switching means can accommodate a maximum of four gear settings per revolution, which is why additional control means are provided for higher gear numbers.

Uncoupling is in accordance with this previously known pawl design 2 from DE 10 2013 112 788 B4 made particularly difficult by the fact that the end face of the locking lever and the coupling surface on the internal teeth of the sun gear form an acute angle α with the perpendicular to the radial locking force, as a result of which the outer edge of the locking lever can get caught in the sun gear under load.

3. Task

It is the object of the invention to create a pawl clutch that considerably simplifies shifting under load and overcomes the disadvantages of the prior art.

4. Summary of the invention

The object is achieved by a pawl clutch according to the invention of a switching device according to the features of independent claim 1. Advantageous embodiments are specified in the dependent claims. Furthermore, a method for dimensioning the pawl design for load-independent switching is specified with dependent claims.

The solution is that the end face of an arm of a locking lever and the corresponding coupling surface of an internal toothing are inclined at a uniform angle with the perpendicular to the direction of the locking force so that there is no counteracting or catching, but on the contrary, a resulting force component when activating a decoupling in the separation direction, which largely compensates for the frictional forces that occur when shifting under load and thereby supports the shifting process.

A pawl clutch of a switching device consists of a pawl mounted on a drive axle with a pivot pin, the end face of which, in the coupled state, engages in a locking manner in a toothing, in particular internal toothing, of a gear wheel, for which purpose it is usually influenced by a spring force is applied. For decoupling, a cam that can be pivoted by adjusting means is arranged on the pawl in such a way that it pivots, if necessary, counteracting a spring force as well as frictional forces that occur in the pawl bearing and between the pawl end face and the teeth, in particular when there is a load caused by a through the pawl transmitted torque occurs, overcomes.

The pawl clutch of a switching device according to the invention comprises a pawl with a locking lever and a control lever, which is pivotally mounted on a pivot pin on a drive axle and in which, when the control lever is pivoted into a cam valley of a cam body, an arm of the locking lever with an end face in one The toothing of a gear wheel on a dome surface engages in a force-fitting manner depending on the load and whenever the control lever comes into contact with a cam due to an adjustment of the cam body, the locking lever releases against a load-dependent frictional force from the internal toothing on the gear wheel, the end face and the dome surface being at an angle β to one Vertical is inclined by a vector extending radially from the pawl pivot axis and leading to the center of the end face, the angle β being inclined such that the radius r _K1 from the pawl pivot axis to the edge of the end face oriented towards the internal toothing is smaller than the radius r _K2 from the pawl pivot axis to the edge of the end face oriented towards the cam body, whereby a load-dependent decoupling force component also occurs against the frictional force.

According to the invention, the load-dependent frictional forces are compensated for by such an inclination of the pawl end faces to the perpendicular to the radial locking force by dividing the force vector that causes the frictional forces according to the respective friction coefficients of the material pairings of the friction surfaces in such a way that, in addition to a coupling force component, there is also a force component in the Separating direction of the coupling arises. As a result, the requirements for the control means for clutch actuation are advantageously largely independent of the load to be switched. This enables a simple and largely load-independent switching process under load. This advantage applies to both pedal and, in particular, motor-driven bicycle transmissions: the drive does not have to be specifically suspended during the switching process in order to enable decoupling.

The person skilled in the art will recognize that, depending on the application, the gear wheel forms either the sun gear with internal teeth of a planetary gear or a gear with internal teeth of a spur gear. The pawl clutch of a switching device and the dimensioning method can therefore be used for different applications with regard to the gear ratio, the required torque and the dimensions of a gearbox on a bicycle or e-bike for shifting under load.

In the following, the application of the gear wheel as a sun gear with internal teeth of a planetary gear is preferably described. The sun gear is rotated by at least three planetary gears. Planetary gears enable an advantageously compact and weight-saving implementation of torque transmission in the drive hub of a bicycle or e-bike.

When using a gear with internal teeth of a spur gear, the gear engages with another gear. In applications where a spur gear is preferred for technical reasons, the advantages of shifting under load according to the invention are easily transferable.

It is advantageous if the resulting decoupling force component corresponds at least to the clutch friction force within narrow tolerances.

Preferably, the angle β is designed such that the resulting force vector with the average distance of the end face from the pivot axis forms a coupling-out moment, which is the sum of the moments from a frictional force F _K with the distance r _K and a bearing friction of the pawl on the pivot pin (with the bearing radius) with tight tolerances.

In the coupled state, the predominance of the uncoupling moment over the frictional moments can be exceeded by a suitably dimensioned spring force of a spring acting in the coupling direction.

Preferably, the locking lever carries at least one locking cam, which touches a cam track, from which a cam/slide track protrudes, which, in a coupling rotational position of the cam body, moves the locking lever with its locking surface against the dome surface, the control lever sliding off the cam into a cam valley .

Advantageously, two locking cams can be formed on the locking lever parallel to one another, positioned on both sides with play next to the cam.

It is also advantageous that the end face and the coupling surface are each designed in an arc shape around the pivot axis.

It is preferred that the cam valley and also the tooth areas of the internal toothing each extend over 30° to 45° of the cam body circumference or the gear wheel interior.

In a preferred embodiment, the toothing can be formed laterally on a gear wheel and the pawl and the cam body can be arranged laterally therefrom, the pivot pin being arranged on a driven gear wheel parallel to the first-mentioned gear wheel, so that both gear wheels rotate together when coupled.

• - Bestimmung des Reibungskoeffizienten zwischen einer Stirnfläche und einer komplementären Kupplungsfläche durch Messung am Objekt und/oder aus einer Tabelle mit Materialeigenschaften,
• - Bestimmung eines Reibungskoeffizienten und eines Radius eines Schwenkzapfens durch Messung am Objekt und/oder aus einer Tabelle mit Materialeigenschaften,
• - Bestimmung eines Abstands der Stirnfläche und der komplementären Kupplungsfläche zur Schwenkachse,
• - Bestimmung eines Winkels β, wobei für die Stirnfläche eines Sperrhebels einer Klinke der Winkel β des stumpfen Winkels 90°+β, der sich gekuppelt formschlüssig in der komplementär ausgebildeten Verzahnung befindet, nur aus solchen Konstruktionsdaten bestimmt wird, die lastunabhängig sind, nämlich dem Reibungskoeffizienten zwischen der Stirnfläche und der komplementären Kupplungsfläche und deren Abstand zur Schwenkachse sowie dem Reibungskoeffizienten und dem Radius des Schwenkzapfens, wodurch sich der Winkel β = arctan((µ_K*r_K + µ_S*r_S) / (r_K - µ_S*µ_K*r_S)) als optimal ergibt.

Furthermore, the present invention solves the problem by means of a method for dimensioning a pawl coupling according to the invention with the following steps:

• - Determination of the coefficient of friction between an end face and a complementary coupling surface by measuring the object and/or from a table with material properties,
• - Determination of a coefficient of friction and a radius of a pivot pin by measuring the object and/or from a table with material properties,
• - Determination of a distance between the end face and the complementary coupling surface to the pivot axis,
• - Determination of an angle β, whereby for the end face of a locking lever of a pawl, the angle β of the obtuse angle 90°+β, which is coupled in a form-fitting manner in the complementary toothing, is determined only from those design data that are load-independent, namely the coefficient of friction between the end face and the complementary coupling surface and their distance to the pivot axis as well as the coefficient of friction and the radius of the pivot pin, whereby the angle β = arctan((µ _K *r _K + µ _S *r _S ) / (r _K - µ _S * µ _K *r _S )) results as optimal.

An approximation for dimensioning can be used by neglecting the term (- µ _S *µ _K *r _S ) in the divisor, whereby the simplified calculation approximates an angle β∼ corresponding to β∼ = arctan (µK + µS *RS/rK) results.

The dimensioning of the angle β of the inclined position of the pawl end face depends on both the dimensions of the pawl and the friction coefficients of the material pairings of the rubbing surfaces on the object. There are tables with practical empirical values of the coefficients of friction, but they show a wide spread of the coefficients. The use of lubricants is also important. It is therefore advisable in individual cases to determine the coefficients of friction on the object using a standard method.

For the following examples, the material pairings are assumed to be steel on steel, lubricated, and the coefficients of friction are set to 0.12, corresponding to a commonly used average value. With the comparative dimensions that a product has according to the DE 10 2013 112 788 B4 have been removed, contrary to the acute angle α used there and counteracting the decoupling, this results in a mathematically optimal inclination of the pawl face normal of 7° to the radial vector through the pawl pivot axis. In another example of the combination of materials and/or lubrication, the calculation resulted in 10° for the inclination of the specified characteristics of the loaded pawl.

Depending on the material pairing and, if necessary, lubrication, the angle β is preferably in a range from 3° to 30°, advantageously between 5° and 20°, preferably between 6° and 12°.

A coefficient of friction depends on the material pairing, the heat treatment such as hardening, lubrication, the contact time, etc. This means that a significant spread in the coefficients of friction is taken into account when determining the angle β.

The preferred material pairings used are steel and steel or steel and aluminum. But there are also other materials in combination with steel or in combination with each other, such as aluminum bronze, bronze, beryllium copper, brass, Teflon, industrial ceramics or solid or Surface-coated components with a low-friction but tribologically resistant surface, for example made of carbon or carbon fibers, are considered in order to be able to realize an angle β at which safe coupling and uncoupling can be guaranteed even under high loads and strong external vibrations. The coefficient of friction of the individual material combinations is determined by measuring the object and/or from a table with material properties.

The determination of the technically different friction coefficients of the respective material pairing and their friction parameters is an integral part of the dimensioning process and determines the angle β resulting in each case according to the invention.

This offers a solution in which the angle β, on the one hand, supports the switching process of decoupling and, on the other hand, enables the clutch to be held securely, instead of, as was previously the case, counteracting a process of decoupling in order to achieve holding the clutch.

5. Brief description of the drawings

• 1 - Stand der Technik aus EP 0 915 800 B1 .
• 2 - Stand der Technik aus DE 10 2013 112 788 B4 .
• 3 - Ein bevorzugtes Ausführungsbeispiel der vorliegenden Erfindung umfassend einen Sperrhebel einer Klinke.
• 4 - Ein bevorzugtes Ausführungsbeispiel der vorliegenden Erfindung einer Klinkenschaltung.
• 5 - Einen Ausschnitt mit formschlüssig gesteuerter Klinke.
• 6 - Eine formschlüssige Klinkenanordnung in perspektivischer Ansicht.
• 7 - Einen Schwenkzapfen mit Sperrhebel und Steuerhebel und einen Nockenkörper in perspektivischer Ansicht.

Exemplary embodiments of the invention are described below with reference to figures. These figures show:

• 1 - State of the art EP 0 915 800 B1 .
• 2 - State of the art DE 10 2013 112 788 B4 .
• 3 - A preferred embodiment of the present invention comprising a locking lever of a pawl.
• 4 - A preferred embodiment of the present invention of a jack circuit.
• 5 - A cutout with a positively controlled latch.
• 6 - A positive latch arrangement in perspective view.
• 7 - A pivot pin with locking lever and control lever and a cam body in a perspective view.

6. Detailed description

Embodiments of the present invention are described below only as examples. These examples represent the best ways of putting the invention into practice currently known to the applicant, although of course these are not the only ways in which this could be accomplished. The description sets forth the functions of the example and the sequence of steps for designing and operating the example. However, the same or equivalent functions and sequences may be achieved from other examples.

The same components have the same reference numbers.

1 shows from the EP 0 915 800 B1 a pawl clutch of a planetary gear in the coupled state. On the hollow drive axle 6, the two-armed pawl 1 is mounted with the pivot pin 2 so that it can pivot about the central pivot axis 8 of the pivot pin 2. One arm of the pawl 1 is the control lever 4, which is pressed into a cam valley NT of a cam body 5 by a spring force F. (The spring force F is not shown because it is familiar to those skilled in the art.) The opposite pawl arm is the locking lever 3, the end face SF of which positively contacts a coupling surface KF of an internal toothing ZI of the sun gear 7. The end faces and coupling surfaces are perpendicular to a force vector which, in the event of a load, is directed from the pivot axis 8 to the end face SF. If you turn the cam body clockwise in the picture, the cam N pushes the control lever 4 into the area of the drive axle 6, whereby it has to overcome the spring force F and the load-dependent friction forces on the end face SF and on the friction surface RF on the pivot pin 2.

2 shows from the DE 10 2013 112 788 the analogous relationships 1 in the representation of a gear wheel. This can either form the sun gear of a planetary gear or a gear of a spur gear. A significant difference to EP 0 915 800 B1 is that the end face SF of the locking lever 3 is not vertical, but at an acute angle α opposite to the vertical to the radial vector which runs through the pawl pivot axis 8, which is inclined towards the interior of the gear wheel in the locking direction and the resulting increased adhesive forces overcome when uncoupling Need to become. For this purpose, a different cam profile is designed, through which the effect of the increased load-dependent frictional forces when releasing the clutch is gradually reduced over a very wide cam twist angle.

3 shows a top view of the new locking lever 3 of the pawl in detail with dimension names and force vectors using the example of the application as a sun gear in a planetary gear. The locking lever 3 can be pivoted in the pivot pin 2 about the pawl pivot axis 8. A bearing in which the pivot pin 2 rotates about the pawl pivot axis 8 is in 3 not shown. The pawl is latched into the sun gear 7 in a form-fitting manner. When force is transmitted to the coupling surface KF of the sun gear 7, a force vector R _K emanating radially from the pawl pivot axis 8 is present on the end face SF of the locking lever 3. The end face SF is not perpendicular to it as in the known arrangements or as in the DE 10 2013 112 788 B4 inclined by an angle in the locking direction, but according to the innovation by an angle β in the opposite direction, i.e. inclined towards the sun gear against the locking direction in such a way that decoupling is supported. Due to its inclination, a normal force N _K is applied perpendicular to the end face SF, which is vectorially composed of the radial force vector R _K and a tangential force vector T _K. It holds for R _K = N _K ∗ cos(β) and for T _K = N _K ∗ sin(β). The load-dependent tangential force T _K is oriented in the direction in which it supports the release of the pawl. On the other hand, the load-dependent frictional force F _K , which occurs between the end face SF and the coupling surface KF on the sun gear 7, as well as the force F _S , which occurs on the friction surface RF on the pivot pin 2 when the pawl is released, act in the opposite direction to the tangential force T _K. According to the invention, this results in extensive compensation of the load-dependent moments M _GES (not in 3 shown). It should be noted that on the friction surface RF of the pivot pin 2, both a proportion of the friction force is determined by the radial force R _K and a smaller further contribution is determined by the friction force on the end face SF proportional to the product of the friction coefficients µ _K and µ _S (both not in 3 shown).

This creates a pawl clutch of a switching device, in which the pawl 1 with the locking lever 3 and the control lever 4 is pivotally mounted on the pivot pin 2 on the drive axle 6 and each time the control lever 4 is pivoted into a cam valley NT of the cam body 5 Arm of the locking lever 3 with the end face SF engages in an internal toothing ZI of the gear wheel on the coupling surface KF in a load-dependent, non-positively coupling manner and whenever the control lever 4 comes into contact with a cam N due to an adjustment of the cam body 5, the locking lever 3 acts against a load-dependent frictional force F _K the internal toothing ZI solves, the end face SF and the dome surface KF being inclined at the angle β to a perpendicular to a vector emanating radially from the pawl pivot axis 8 and leading to the center of the end face SF, the angle β being inclined such that the Radius r _K1 from the pawl pivot axis 8 to the edge 31 of the end face SF, which is oriented towards the internal toothing ZI, is smaller than the radius r _K2 from the pawl pivot axis 8 to the edge 32 of the end face SF, which is oriented towards the cam body 5, whereby a likewise load-dependent decoupling force component is counteracted the friction force F _K occurs.

4 shows an additionally optimized pawl arrangement using the example of the application as a sun gear in a planetary gear. This arrangement can also be transferred, for example, to the gear of a spur gear. On the one hand, the end face SFR of the locking lever 3 is designed to be rounded circularly around the pawl pivot axis 8; Likewise, the complementary gear flank of the internal toothing ZI of the sun gear 7 is designed as a rounded coupling surface KFR, which ensures both a stop in the engaged state and further facilitates uncoupling. Accordingly, the control lever 4 is rounded and the cam profile in the cam valley NT and its outer edges are also rounded. The switching angle that the entire cam valley NT requires is less than 45° and is therefore advantageously compared to the switching angle specified at approximately 100° in 2 The control curves shown are far less than half. Therefore, according to the invention, more switching positions can be arranged on one revolution of cam bodies lined up in parallel without critical overlaps occurring during switching.

The angle β of a rounded end face SFR of the locking lever is determined with respect to the connecting line of the corner points of the rounding. The rounding results in an even distribution of the locking force across the pair of friction surfaces, which means that defined friction conditions exist when decoupling. In order to create the most balanced compensation possible for the load-dependent forces, a new method is used, the mathematical derivation of which is presented in detail. The equations given are numbered between slashes. The method is based on the objective that the load-dependent forces acting on the pawl that occur during uncoupling should largely compensate for each other, that is, their total moment M _GES is essentially zero. This leads to equation /5/. The friction forces are each per proportional to the associated friction coefficient µ _K or µ _S. The friction force F _K on the coupling surface SF or SFR results from the normal force N _K , see /2/. The lengths of the respective force arms specified in the moment equations are the radius rs from the pivot pin 2 or the radial distance r _K of the coupling surface KF or KFR from the pawl pivot axis 8. Solve the equation /5/ of the total moment M _GES according to the requirement Equation /6/, an optimized angle β results from equation /14/. By neglecting a term that is relatively small in practical terms, this results in equation /15/, which establishes a connection between the approximate angle β∼ from the two friction coefficients mentioned and the ratio of the radii of the friction surfaces to the pivot axis 8.

The following mathematical formulas illustrate the findings according to the invention regarding an advantageous dimensioning of the shape of the locking lever and its complementary coupling surface:

The normal force N _K acts on the end face SF of the pawl 1 to a virtually fixed sun gear 7. This can be divided into a component R _K radial for supporting the pawl N _K *cos(ß) and a tangential component T _K. The following applies: $${\text{T}}_{\text{K}}={\text{N}}_{\text{K}}*\text{sin}\left(\mathrm{\beta }\right)$$

The friction force F _K acts on the end face SF, with: $${\text{F}}_{\text{K}}={\mathrm{\mu }}_{\text{K}}*{\text{N}}_{\text{K}}$$

This also has a radial component F _K *sin(β). The overall effect is the radial force A _K : $$\begin{array}{l}{\text{A}}_{\text{K}}={\text{N}}_{\text{K}}*\text{cos}\left(\mathrm{\beta }\right)+{\text{F}}_{\text{K}}*\text{sin}\left(\mathrm{\beta }\right)={\text{N}}_{\text{K}}*\text{cos}\left(\mathrm{\beta }\right)+{\mathrm{\mu }}_{\text{K}}*{\text{N}}_{\text{K}}*\text{sin}\left(\mathrm{\beta }\right)=\\ {\text{N}}_{\text{K}}*\left(\text{cos}\left(\mathrm{\beta }\right)+{\mathrm{\mu }}_{\text{K}}*\text{sin}\left(\mathrm{\beta }\right)\right)\end{array}$$

The normal force Ns on the pivot pin 2 is equal to A _K in magnitude. The friction force F _S is therefore given by: $${\text{F}}_{\text{S}}={\mathrm{\mu }}_{\text{S}}*{\text{N}}_{\text{S}}={\mathrm{\mu }}_{\text{S}}*{\text{A}}_{\text{K}}={\mathrm{\mu }}_{\text{S}}*{\text{N}}_{\text{K}}*\left(\text{cos}\left(\mathrm{\beta }\right)+{\mathrm{\mu }}_{\text{K}}*\left(\mathrm{\beta }\right)\right)$$

The total torque M _GES acting on the pawl 1 has three components, generated by the forces T _K , F _K and Fs, with the associated radii r _K and r _S : $${\text{M}}_{\text{TOTAL}}={\text{T}}_{\text{K}}*{\text{r}}_{\text{K}}-{\text{F}}_{\text{K}}*\text{cos}\left(\mathrm{\beta }\right)*{\text{r}}_{\text{K}}-{\text{F}}_{\text{S}}*{\text{r}}_{\text{S}}$$

$$<=>{\text{N}}_{\text{K}}*\text{sin}\left(\mathrm{\beta }\right)*{\text{r}}_{\text{K}}-{\mathrm{\mu }}_{\text{K}}*{\text{N}}_{\text{K}}*\text{cos}\left(\mathrm{\beta }\right)*{\text{r}}_{\text{K}}-{\mathrm{\mu }}_{\text{S}}*{\text{N}}_{\text{K}}*\left(\text{cos}\left(\mathrm{\beta }\right)+{\mathrm{\mu }}_{\text{K}}*\text{sin}\left(\mathrm{\beta }\right)\right)*{\text{r}}_{\text{S}}=\text{o}$$

$$<=>\text{sin}\left(\mathrm{\beta }\right)*{\text{r}}_{\text{K}}-{\mathrm{\mu }}_{\text{K}}*\text{cos}\left(\mathrm{\beta }\right)*{\text{r}}_{\text{K}}-{\mathrm{\mu }}_{\text{S}}*\left(\text{cos}\left(\mathrm{\beta }\right)+{\mathrm{\mu }}_{\text{K}}*\text{sin}\left(\mathrm{\beta }\right)\right)*{\text{r}}_{\text{S}}=\text{o}$$

$$<=>\text{sin}\left(\mathrm{\beta }\right)*{\text{r}}_{\text{K}}-{\mathrm{\mu }}_{\text{K}}*\text{cos}\left(\mathrm{\beta }\right)*{\text{r}}_{\text{K}}-{\mathrm{\mu }}_{\text{S}}*\text{cos}\left(\mathrm{\beta }\right)*{\text{r}}_{\text{S}}+{\mathrm{\mu }}_{\text{S}}*{\mathrm{\mu }}_{\text{K}}*\text{sin}\left(\mathrm{\beta }\right)*{\text{r}}_{\text{S}}=\text{o}$$

$$\begin{array}{cc}<=>& \text{sin}\left(\mathrm{\beta }\right)*{\text{r}}_{\text{K}}-{\mathrm{\mu }}_{\text{s}}{\mathrm{\mu }}_{\text{K}}*\text{sin}\left(\mathrm{\beta }\right)*{\text{r}}_{\text{S}}={\mathrm{\mu }}_{\text{K}}*\text{cos}\left(\mathrm{\beta }\right)*{\text{r}}_{\text{K}}+{\mathrm{\mu }}_{\text{S}}*\text{cos}\left(\mathrm{\beta }\right)*{\text{r}}_{\text{S}}\end{array}$$

$$\begin{array}{cc}<=>& \text{sin}\left(\mathrm{\beta }\right)*\left({\text{r}}_{\text{K}}-{\mathrm{\mu }}_{\text{S}}*{\text{r}}_{\text{S}}\right)=\text{cos}\left(\mathrm{\beta }\right)*\left({\mathrm{\mu }}_{\text{K}}*{\text{r}}_{\text{K}}+{\mathrm{\mu }}_{\text{S}}*{\text{r}}_{\text{S}}\right)\end{array}$$

$$\begin{array}{cc}<=>& \text{sin}\left(\mathrm{\beta }\right)/\text{cos}\left(\mathrm{\beta }\right)=\left({\mathrm{\mu }}_{\text{K}}*{\text{r}}_{\text{K}}+{\mathrm{\mu }}_{\text{S}}*{\text{r}}_{\text{S}}\right)/\left({\text{r}}_{\text{K}}-{\mathrm{\mu }}_{\text{S}}*{\mathrm{\mu }}_{\text{K}}*{\text{r}}_{\text{S}}\right)\end{array}$$

$$\begin{array}{cc}<=>& \text{tan}\left(\mathrm{\beta }\right)=\left({\mathrm{\mu }}_{\text{K}}*{\text{r}}_{\text{K}}+{\mathrm{\mu }}_{\text{S}}*{\text{r}}_{\text{S}}\right)\end{array}/\left({\text{r}}_{\text{K}}-{\mathrm{\mu }}_{\text{S}}*{\mathrm{\mu }}_{\text{K}}*{\text{r}}_{\text{S}}\right)$$

$$\begin{array}{cc}<=>& \mathrm{\beta }=\text{arctan }\left(\left({\mathrm{\mu }}_{\text{K}}*{\text{r}}_{\text{K}}+{\mathrm{\mu }}_{\text{S}}*{\text{r}}_{\text{S}}\right)/\left({\text{r}}_{\text{K}}-{\mathrm{\mu }}_{\text{S}}*{\text{r}}_{\text{K}}*{\text{r}}_{\text{K}}\right)\right)\end{array}$$

In the case that the friction forces are exactly compensated, M _GES = 0, from which β can be determined as follows: $${\text{T}}_{\text{K}}*{\text{r}}_{\text{K}}-{\text{F}}_{\text{K}}*\text{cos}\left(\mathrm{\beta }\right)*{\text{r}}_{\text{K}}-{\text{F}}_{\text{S}}*{\text{r}}_{\text{S}}=0$$

$$<=>{\text{N}}_{\text{K}}*\text{sin}\left(\mathrm{\beta }\right)*{\text{r}}_{\text{K}}-{\mathrm{\mu }}_{\text{K}}*{\text{N}}_{\text{K}}*\text{cos}\left(\mathrm{\beta }\right)*{\text{r}}_{\text{K}}-{\mathrm{\mu }}_{\text{S}}*{\text{N}}_{\text{K}}*\left(\text{cos}\left(\mathrm{\beta }\right)+{\mathrm{\mu }}_{\text{K}}*\text{sin}\left(\mathrm{\beta }\right)\right)*{\text{r}}_{\text{S}}=\text{O}$$

$$<=>\text{sin}\left(\mathrm{\beta }\right)*{\text{r}}_{\text{K}}-{\mathrm{\mu }}_{\text{K}}*\text{cos}\left(\mathrm{\beta }\right)*{\text{r}}_{\text{K}}-{\mathrm{\mu }}_{\text{S}}*\left(\text{cos}\left(\mathrm{\beta }\right)+{\mathrm{\mu }}_{\text{K}}*\text{sin}\left(\mathrm{\beta }\right)\right)*{\text{r}}_{\text{S}}=\text{O}$$

$$<=>\text{sin}\left(\mathrm{\beta }\right)*{\text{r}}_{\text{K}}-{\mathrm{\mu }}_{\text{K}}*\text{cos}\left(\mathrm{\beta }\right)*{\text{r}}_{\text{K}}-{\mathrm{\mu }}_{\text{S}}*\text{cos}\left(\mathrm{\beta }\right)*{\text{r}}_{\text{S}}+{\mathrm{\mu }}_{\text{S}}*{\mathrm{\mu }}_{\text{K}}*\text{sin}\left(\mathrm{\beta }\right)*{\text{r}}_{\text{S}}=\text{O}$$

$$\begin{array}{cc}<=>& \text{sin}\left(\mathrm{\beta }\right)*{\text{r}}_{\text{K}}-{\mathrm{\mu }}_{\text{s}}{\mathrm{\mu }}_{\text{K}}*\text{sin}\left(\mathrm{\beta }\right)*{\text{r}}_{\text{S}}={\mathrm{\mu }}_{\text{K}}*\text{cos}\left(\mathrm{\beta }\right)*{\text{r}}_{\text{K}}+{\mathrm{\mu }}_{\text{S}}*\text{cos}\left(\mathrm{\beta }\right)*{\text{r}}_{\text{S}}\end{array}$$

$$\begin{array}{cc}<=>& \text{sin}\left(\mathrm{\beta }\right)*\left({\text{r}}_{\text{K}}-{\mathrm{\mu }}_{\text{S}}*{\text{r}}_{\text{S}}\right)=\text{cos}\left(\mathrm{\beta }\right)*\left({\mathrm{\mu }}_{\text{K}}*{\text{r}}_{\text{K}}+{\mathrm{\mu }}_{\text{S}}*{\text{r}}_{\text{S}}\right)\end{array}$$

$$\begin{array}{cc}<=>& \text{sin}\left(\mathrm{\beta }\right)/\text{cos}\left(\mathrm{\beta }\right)=\left({\mathrm{\mu }}_{\text{K}}*{\text{r}}_{\text{K}}+{\mathrm{\mu }}_{\text{S}}*{\text{r}}_{\text{S}}\right)/\left({\text{r}}_{\text{K}}-{\mathrm{\mu }}_{\text{S}}*{\mathrm{\mu }}_{\text{K}}*{\text{r}}_{\text{S}}\right)\end{array}$$

$$\begin{array}{cc}<=>& \text{tan}\left(\mathrm{\beta }\right)=\left({\mathrm{\mu }}_{\text{K}}*{\text{r}}_{\text{K}}+{\mathrm{\mu }}_{\text{S}}*{\text{r}}_{\text{S}}\right)\end{array}/\left({\text{r}}_{\text{K}}-{\mathrm{\mu }}_{\text{S}}*{\mathrm{\mu }}_{\text{K}}*{\text{r}}_{\text{S}}\right)$$

$$\begin{array}{cc}<=>& \mathrm{\beta }=\text{arctan}\left(\left({\mathrm{\mu }}_{\text{K}}*{\text{r}}_{\text{K}}+{\mathrm{\mu }}_{\text{S}}*{\text{r}}_{\text{S}}\right)/\left({\text{r}}_{\text{K}}-{\mathrm{\mu }}_{\text{S}}*{\text{r}}_{\text{K}}*{\text{r}}_{\text{K}}\right)\right)\end{array}$$

Since r _K is practically considerably larger than µ _S *µ _K *r _S , the latter can be neglected to obtain the following approximation: $$\mathrm{\beta }\sim \mathrm{=}\text{arctan}\left(\left({\mathrm{\mu }}_{\text{K}}*{\text{r}}_{\text{K}}+{\mathrm{\mu }}_{\text{K}}*{\text{r}}_{\text{K}}\right)/{\text{r}}_{\text{K}}\right)=\text{arctan}\left({\mathrm{\mu }}_{\text{K}}+{\mathrm{\mu }}_{\text{S}}*{\text{r}}_{\text{S}}/{\text{r}}_{\text{K}}\right)$$

Accordingly, an approximation for dimensioning can be used by neglecting a defined part of the term, namely the term (- µ _S *µ _K *r _S ) in the divisor, whereby the simplified calculation uses an angle (β∼) corresponding to β as an approximation ∼ = arctan (µ _K + µ _S *R _S /r _K ) results.

If you choose example values for the radii and coefficients of friction that known objects have, the design method shown leads to the following result as an example: $${\text{r}}_{\text{S}}=3\text{mm},{\text{r}}_{\text{K}}=6.5\text{mm},{\mathrm{\mu }}_{\text{K}}={\mathrm{\mu }}_{\text{S}}=0.12->\mathrm{\beta }=10.01°,\mathrm{\beta }\sim =\text{9},\text{95}°$$

It turns out that the approximation according to equation /15/ only deviates by less than 1% from the exact value according to equation /14/. The tolerances for determining the coefficients of friction and their long-term behavior are considerably higher. Therefore, it must be decided in each case how the tolerance field should be set between over- and under-compensation. In the event of overcompensation, a sufficient spring force reserve must be provided to keep the clutch securely closed.

The illustrated type of switchable clutch of a gear wheel with a drive element by means of a cam-controlled pawl, which is arranged inside a sun gear, for example, as is usual in the prior art, can be easily, with all the innovations, based on a lateral offset of the pawl with the cam body and the Transmit drive element if the gear wheel has lateral clutch teeth. If several gears are arranged side by side [as is usual in the prior art], the pivot pin of the pawl can be mounted on an adjacent driving gear, whereby the respective coupled gears rotate together.

A further advantageous embodiment of a coupling device with a pawl is based on 5 and 6 described. Here, the coupling and uncoupling of the pawl is made possible without spring force by adjusting the pawl in the coupling or uncoupling direction in a form-fitting manner each time the cam body is rotated by means of additional cams, with the decoupling inclination of the end face of the locking lever described above also being used becomes. Significant tolerances in the coefficients of friction can therefore be easily adapted.

5 shows a positively controlled pawl arrangement in the decoupled position using the example of a planetary gear. This arrangement can also be transferred to the gear of a spur gear. The control lever 4 is located at the height of the cam N, whereby the rounded locking surface SFR on the locking lever 3 is separated from the rounded coupling surface KFR and the locking lever 3 can freely pass the teeth of the gear wheel, namely the sun wheel, until the cam body 5 is in the direction of the arrow has been twisted far enough that a projection G ₂ on a skid has passed under a locking cam S ₂ , which is formed on the locking lever 3, and pushes it out so far that the locking surface SFR contacts the coupling surface KFR and thus the coupled state, not shown here manufactures. A spring force is not provided for coupling in this exemplary embodiment, and an overcompensated clutch separating force, which can occur, for example, when the coefficient of friction is reduced, has no disruptive consequences for load transfer when a coupling from the pivot pin 2 to the sun gear 7 takes place.

6 shows a perspective detail of an advantageous latch design according to. 5 in the coupled position. The pivot pin 2, which is not visibly connected to a hollow drive axle 6, carries the previously described locking and control levers 3, 4, which interact with the cam N on the cam body 5 to release the clutch when the cam body is rotated. The locking lever 3 now has two parallel locking cams S ₂ , S ₃ , which interact in a coupling manner with projections G ₂ , G ₃ on slideways of the cam body 5 when it is in the position shown. The entire area from the start of engaging to the end of disengaging during continuous switching requires an approximately 30° rotation of the cam body 5, so that in principle 12 different settings can be accommodated on the entire circumference of the cam body.

The positive control of the pawl clutch has the further advantage that springs for pawl actuation can be omitted and the elements that effect the new control, namely the sliding tracks with their elevations and the locking cams are each manufactured completely integrated with the cam body or the locking lever.

The pairwise arrangement of the slideways and the locking cams on both sides of the reverse-acting cam N advantageously avoids the creation of torsional forces on the locking lever 3 as well as bending forces on the pivot pin with possible increased friction and wear in the pivot bearing.

Alternatively, two parallel control levers are provided over aligned parallel cams, and only one locking cam is arranged with an associated sliding cam between the other two cams with sufficient play.

7 shows a pivot pin with locking lever and control lever and a cam body in a perspective view.

The pivot pin 2 with the control lever 4 and the locking lever 3 is rotatably mounted about a pawl pivot axis 8. The warehouse is not shown. Starting from the pawl pivot axis 8, the radius r _K1 is formed towards the edge 31 of the end face SF on the locking lever 3, which is oriented towards the internal toothing ZI (not shown). Starting from the pawl pivot axis 8, the radius r _K2 is formed towards the edge 32 of the end face SF on the locking lever 3, which is oriented towards the cam body 5.

1

Klinke

2

Schwenkzapfen

RF

Reibfläche an 2

3

Sperrhebel

31

zur Innenverzahnung ZI orientierte Kante

32

zum Nockenkörper orientierte Kante

SF

Stirnfläche an 3

SFR

Stirnfläche gerundet

S2, S3

Sperrnocken an 3

4

Steuerhebel

5

Nockenkörper

N

Nocke

NT

Nockental in 5

NP

Nockenprofil

G2, G3

Gleitbahnen

F

Federkraft

6

Antriebsachse

7

Sonnenrad

ZI

Innenverzahnung

KF

Kupplungsfläche

KFR

Kupplungsfläche gerundet

8

Klinkenschwenkachse

NK

Normalkraft an SF

FK

Reibkraft an SF

RK

Radialkomponente von N_K

TK

Tangentialkomponente von N_K

µK

Reibungskoeffizient der Kupplungsflächen SF zu KF

AK

Radialkraft in 1

NS

Normalkraft an RF

FS

Reibkraft an RF

V

Drehrichtung

µS

Reibungskoeffizient an RF

MGES

Gesamtmoment

rS

Radius von 2

rK

Abstand KF von 8

α

spitzer Winkel

β

Winkel

β∼

Winkel genähert

The list of reference symbols refers to the accompanying figures and the method of dimensioning.

1

pawl

2

Pivot pin

RF

Friction surface on 2

3

Locking lever

31

Edge oriented towards the internal toothing ZI

32

Edge oriented towards the cam body

SF

Face on 3

SFR

Rounded end face

S2, S3

Locking cam on 3

4

Control lever

5

cam body

N

cam

NT

Cam valley in 5

NP

Cam profile

G2, G3

Slideways

F

Spring force

6

Drive axle

7

sun gear

ZI

Internal gearing

KF

Coupling surface

KFR

Coupling surface rounded

8th

Pawl pivot axis

NK

Normal force at SF

FK

Friction force on SF

RK

Radial component of N _K

TK

Tangent component of N _K

µK

Friction coefficient of the coupling surfaces SF to KF

AK

Radial force in 1

NS

Normal force at RF

FS

Friction force on RF

v

Direction of rotation

µS

Coefficient of friction at RF

MGES

Total moment

RS

Radius of 2

rK

Distance KF of 8

α

acute angle

β

angle

β∼

Angle approached

QUOTES INCLUDED IN THE DESCRIPTION

This list of documents listed by the applicant was generated automatically and is included solely for the better information of the reader. The list is not part of the German patent or utility model application. The DPMA assumes no liability for any errors or omissions.

Cited patent literature

• EP 0915800 B1 [0002, 0032, 0035, 0036]
• DE 102013112788 B4 [0003, 0005, 0026, 0032, 0037]
• DE 102013112788 [0036]

Claims (14)

Pawl clutch of a switching device, in which a pawl (1) with a locking lever (3) and a control lever (4) is pivotally mounted on a pivot pin (2) on a drive axle (6) and each time the control lever (4) is in a cam valley (NT) of a cam body (5) is pivoted in, an arm of the locking lever (3) with an end face (SF) engages in a force-fitting, load-dependent manner in an internal toothing (ZI) of a gear wheel on a coupling surface (KF) and in each case when by adjusting the Cam body (5) the control lever (4) comes into contact with a cam (N) the locking lever (3) releases from the internal toothing (ZI) against a load-dependent frictional force (F _K ), characterized by the end face (SF) and the coupling surface (KF) is inclined at an angle (β) to a perpendicular to a vector extending radially from the pawl pivot axis (8) and leading to the center of the end face (SF), the angle (β) being inclined such that the radius (r _K1 ) from the pawl pivot axis (8) to the edge (31) of the end face (SF) oriented towards the internal toothing (ZI) is smaller than the radius (r _K2 ) from the pawl pivot axis (8) to the edge (31) oriented towards the cam body (5) ( 32) of the end face (SF), whereby a load-dependent decoupling force component occurs against the friction force (F _K ). Pawl clutch Claim 1 , characterized in that the angle (β) is designed such that the resulting decoupling force component corresponds at least to the clutch friction force (F _K ) within narrow tolerances. Pawl clutch Claim 1 , characterized in that the angle (β) is designed such that the resulting force vector with the average distance (r _K ) of the end face (SF) from the pivot axis (8) forms a coupling-out moment that is the sum of the moments from one Frictional force (F _K ) with the distance (r _K ) and a bearing friction (Fs) of the pawl (1) on the pivot pin (2) with the bearing radius (rs) corresponds to narrow tolerances. Pawl clutch Claim 3 , characterized in that in the coupled state, a predominance of the decoupling moment over the frictional moments is exceeded by a suitably dimensioned spring force (F) of a spring acting in the coupling direction. Pawl clutch according to one of the Claims 1 until 3 , characterized in that the locking lever (3) carries at least one locking cam (S ₂ , S ₃ ), which touches a cam track, from which a cam body (5) protrudes, which in a coupling rotational position of the cam body (5) controls the locking lever ( 3) with its locking surface (SFR) against the coupling surface (KFR), whereby the control lever (4) slides from the cam (N) into a cam valley (NT). Pawl clutch Claim 5 , characterized in that two locking cams (S2, S3) are formed on the locking lever (3) parallel to each other, positioned on both sides with play next to the cam (N). Pawl coupling according to at least one of the preceding claims, characterized in that the end face (SFR) and the coupling surface (KFR) are each designed in an arc shape around the pivot axis (8). Pawl clutch according to one of the preceding claims, characterized in that the cam valley (NT) and also the tooth areas of the internal toothing (ZI) each extend over 30° to 45° of the cam body circumference or the inside of the gear wheel. Pawl clutch Claim 1 , characterized in that the gear wheel forms either a sun gear (7) with internal teeth (ZI) of a planetary gear or a gear with internal teeth (ZI) of a spur gear. Pawl coupling according to one of the preceding claims, characterized in that the toothing is formed laterally on a gear wheel and the pawl (1) and the cam body are arranged laterally therefrom, the pivot pin (2) being arranged on a driven gear wheel parallel to the first-mentioned gear wheel , so that both gear wheels rotate together when coupled. Pawl coupling according to one of the preceding claims, characterized in that the angle β is preferably in a range from 3° to 30°, advantageously between 5° and 20°, in particular between 6° and 12°. Method for dimensioning a pawl clutch according to one of the preceding claims, characterized by the steps - Determination of the coefficient of friction (µ _K ) between an end face (SF, SFR) and a complementary coupling surface (KF, KFR) by measuring on the object and / or from a table with material properties - determination of a coefficient of friction (µ _s ) and a radius (r _s ) of a pivot pin (2) by measurement on the object and/or from a table with material properties - determination of a distance (r _K ) of the end face (SF, SFR) and the complementary coupling surface (KF, KFR) to the pivot axis (8) - determination of an angle (β), whereby for the end face (SF, SFR) of a locking lever (3) of a pawl (1) the angle (β) of the obtuse angle (90 ° + β), which is coupled in a form-fitting manner in the complementary toothing, is only determined from those design data that are load-independent, namely the coefficient of friction (µ _K ) between the end face (SF, SFR) and the complementary coupling surface (KF, KFR ) and their distance (r _K ) to the pivot axis (8) as well as the coefficient of friction (µ _S ) and the radius (rs) of the pivot pin (2), whereby the angle (β) = arctan((µ _K *r _K + µ _S *r _S ) / (r _K - µ _S *µ _K *r _S )) results as optimal. Procedure according to Claim 12 , characterized in that an approximation for dimensioning is used by neglecting the term (- µ _S *µ _K *r _S ), whereby the simplified calculation approximates an angle (β∼) corresponding to β∼ = arctan (µ _K + µ _S *R _S /r _K ) results. Procedure according to Claim 12 or 13 , characterized in that the coefficient of friction for a rubbing steel/steel material pair is preferably set at 0.12 from a table or this is determined in advance by measurements.

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