DE102009015952B4 - Rattle-free component pairing - Google Patents

Abstract

A gear pair (10) having a first gear (12) with a first toothing (16) and a second gear (14) with a second toothing (18) which is in engagement with the first toothing (16) to engage over the toothings (16). 16, 18) to transmit a driving force in a rotational direction (20), wherein the first gear (12) further comprises a first anti-rattle toothing (32), wherein on the second gear (14) a anti-ratchet gear (30) is attached, the a second anti-rattle toothing (33) which is in engagement with the first anti-rattle toothing (32), the anti-rattle toothed wheel (30) being mounted in the radial direction (115) movably with respect to the second toothed wheel (14), a radial stop (50 ) relative movement of the anti-ratchet gear (30) in the radial direction with respect to the second gear (14) limited so that a maximum deflection (60) is established, characterized in that the anti-rattle gear (30) so elastically deformable forms and / or is elastically mounted in the radial direction, ...

Description

The present invention relates to a gear pair of a first gear with a first toothing and a second gear with a second toothing, which is in engagement with the first toothing in order to be able to transmit a driving force in a drive direction via the toothings.

Such gear pairs are often used in drive trains of motor vehicles, for example, in multi-step gearboxes, in drives for auxiliary equipment, etc.

One of the main problems with such gear pairs is the so-called rattle phenomenon. This occurs mainly due to vibration excitations in the drive train, which are generated for example by a drive motor such as an internal combustion engine of the drive train. The rattling (also referred to as unsympathetic vibrations) arises from the fact that due to the vibration excitation delays the driving gear, the driven gear (eg a loose wheel) but continues to rotate with an impressed orbital motion and is delayed only by friction and drag torque effects. In this case, the driven gear from a pull flank of the drive gear releases to swing to the thrust flank of the drive gear and possibly abut there. Such phenomena occur not only in load change reactions, but in particular due to the higher-frequency excitations from other parts of the drive train, such as an internal combustion engine.

To reduce such sounds, there are several approaches. On the one hand active external gear measures can be provided, which decouple, for example, the disturbance excitation from an internal combustion engine by a dual mass flywheel. However, such dual-mass flywheels are complex in terms of the claimed installation space, the necessary additional weight and in terms of costs. Another possibility is passive external gear measures, such as encapsulation or insulation of the gear housing. These measures are unfavorable. Furthermore, active internal gear measures are known, which are arranged specifically to the main noise sources. Such active internal transmission measures are often aimed at minimizing the functional games or hindering the mobility within these functional games. The disadvantage here is often the reduced efficiency and the generation of other unwanted noise (such as howling). Furthermore, it is known to provide noise reduction passive internal gear measures that are located directly on the noise sources (ie, for example, the gears) and erase or isolate mechanical vibrations.

Well-known measures here are Losradbremsen, measures to Zahnlückenverspannung, measures in which a disc is used with a slightly different ratio, measures with a friction wheel Nebenübersetzung, vibration absorber, magnetic solutions to prevent release of the tooth flanks from each other, etc.

For example, from the document DE 103 28 482 A1 a gear transmission with a anti-rattle device known. In this case, a loose wheel and a fixed gear are each assigned a friction wheel, which are in frictional engagement with each other.

From the document DE 197 21 851 A1 It is known to reduce the backlash in a gear pair by a toothed wheel with slightly bendable teeth is attached to the one gear. The deflectable teeth of the toothed wheel engage in the counter toothing of the gear pair and should provide for a noise reduction without significant wear on the other gear member.

From the DE 38 39 807 C1 It is known to cancel the backlash between two gears by an additional toothed disc is provided on a gear and biased by the toothed disc relative to the associated gear by springs in the circumferential direction.

Further, the document discloses DE 10 2004 008 171 A1 a spur gear for camshaft in which a gear is formed in two parts.

Anti-rattle measures are also known from the prior art (eg. DE 1 967 959 U . JP 62228735 A . US 4,577,525 A ), in which an anti-rattle toothing of a anti-ratchet gear has a tooth more or less than the toothing of the associated gear.

The document JP 01153865 A discloses an arrangement having a gear pair and an associated anti-ratchet gear, wherein the ratchet teeth of the gear, with which the anti-ratchet gear is engaged, has a different helix angle than the teeth thereof.

From the US 2003/0136212 A1 a gear pairing is known in which a gear is assigned a anti-ratchet gear. The anti-ratchet gear is supported on the associated gear so as to be frictionally engaged with the gear but movable in the rotational direction and in the radial direction stored for this purpose. Furthermore, the anti-rattle component is fastened by a plurality of screws on the associated gear. In this case, a radial clearance between the anti-rattle gear and the outer peripheral portions of the screws is established.

In the DE 73 45 374 U is described a backlash-free pinion. In this case, there is a wheel of a meshing gear pair of two sprockets, which are biased against each other by a force acting between the two sprockets circumferentially acting spring. One of the sprockets takes over the torque transmission, whereas the other is responsible for the clearance compensation.

Against the above background, it is the object of the invention to provide a gear pairing, with which an effective noise reduction can be realized and has a high efficiency.

According to a first aspect of the present invention, the above object is achieved by a gear pair of a first gear with a first toothing and a second gear with a second toothing, which is in engagement with the first toothing to drive a driving force in a drive direction via the toothings Further, the first gear is associated with a first anti-rattle toothing, wherein on the second gear a anti-ratchet gear is fixed, which has a second anti-rattle toothing, which is in engagement with the first anti-rattle teeth, and in the radial direction (ie substantially perpendicular or transversely to the drive direction) is movable, and so elastically deformable and / or elastically mounted in the radial direction, that it is pressed away in the region of the meshing with the first anti-rattle toothing in the radial direction of the first gear, wherein a radial stop a Relativb Enwegung the anti-ratchet gear in the radial direction with respect to the second gear limited so that a maximum deflection is established, and wherein the radial stop over the circumference of the second gear has a uniform radius.

By means of the radial stop provided according to the invention, it can be ensured that the anti-ratchet toothed wheel is not deflected too eccentrically with respect to the second toothed wheel due to centripetal or centrifugal forces. The reliability of the gear pairing can thus be increased.

The first aspect of the invention preferably consists in that the largest occurring rotational backlash of the teeth can be compensated by the radial deformability or radially elastic mounting of the anti-ratchet gear, since the anti-ratchet toothed wheel can thus deflect radially elastically in the region of the toothed engagement. In other words, the tooth engagement between the anti-rattle toothings is such that such a force acts between the anti-rattle toothed wheel and the second toothed wheel in the drive direction that the toothings do not become so within the range even with high-frequency excitations (such as, for example, from an internal combustion engine, in particular a diesel engine) Backlash game turn that a rattle noise is generated. In other words, a drag torque, in particular a friction torque in the circumferential direction can be generated on the anti-rattle toothing, which damps the disturbing suggestions. The drag torque should be suitable to reduce the oscillating mass forces, in particular to eliminate.

In this embodiment, all production-related tolerances and thermal deformations, which have an effect on the rotational flank play, can be compensated preferably up to 100%.

In this case, the anti-rattle toothed wheel itself can be elastically deformable so that it can yield radially elastically in the region of the toothed engagement. Alternatively or additionally, the anti-rattle gear can be mounted elastically in the radial direction. In this embodiment, due to the radially elastic deflection in the region of the meshing engagement, a center of the anti-ratchet gear may be arranged eccentrically offset from a center of the second gear. Here, in other words, any point of the anti-ratchet gear may make a circular motion around the center of the second gear.

The forces that lead to a radial deflection of the anti-ratchet gear can be returned to the system when leaving the meshing due to the radially elastic property of the radial spring element.

The anti-ratchet gear is arranged or formed so that it is pushed away from the first gear in the radial direction in the region of the tooth engagement with the first ratchet toothing, that is, it deflects radially elastically in the region of the tooth engagement.

In this case, the anti-ratchet gear is usually biased toward the first gear, so that the radial deflection takes place against the bias. In any case, a radial deflection of the anti-ratchet gear in the region of the meshing takes place in any case.

If the anti-ratchet gear and the second gear are connected to each other via a frictional engagement, the frictional force can be increased by this deflection, so that Relative movements between the second gear and the anti-ratchet gear in the drive direction due to the frictional engagement are attenuated more. As a result, a turning of the teeth and thus rattling or rattling can be prevented.

However, the radial deflectability of the anti-ratchet gear with respect to the second gear on which the anti-rattle gear is fixed, not only allows an increase in the frictional forces. Also, clamping effects during a two-flank rolling engagement can be reduced and preferably avoided.

Since the radial pushing away of the anti-rattle gear is done with a relatively small force, the loss of efficiency is substantially negligible. In addition, the forces required for radial deflection at the exit from the tooth engagement can be at least partially returned.

In the gear pairing, the driving direction is a driving direction directed in the circumferential direction of the driving gear. In this case, you can take a gear as a split wheel, with one part is responsible for the load transfer and the other part (anti-rattle gear) can provide the backlash. However, the gear pair may also be a combination of a rack and a gear, wherein the drive direction is substantially linear.

The gear pairing according to the invention is also effective in both drive directions (that is to say in the case of gears, for example in both directions of rotation). Furthermore, the gear pair according to the invention is applicable to both straight and helical gears.

The anti-rattle teeth can be identical or similar in terms of tooth shapes as the teeth. However, the anti-rattle toothings can also have any other shape, it being preferred for the anti-rattle toothing to engage with one another in a punctiform or linear manner. It is particularly preferred if the anti-rattle toothings are punctiform or linearly engaged with each other at the height of the pitch circle.

In the gear pairing according to the invention, it is generally irrelevant whether the first or the second gear is the driving gear. Preferably, however, the anti-rattle gear is connected to a fixed gear (ie, a gear fixedly connected to a rotary shaft) because the idler gear engaged therewith is often installed by clutches (synchronizers, etc.) so that the anti-ratchet gear will not ride there or only with it greater effort can be attached.

Furthermore, it is understood that a gear and two or more anti-rattle gears may be assigned, for example, on axially opposite sides of a gear. This is particularly preferred when the second gear is engaged with more than one first gear. Each anti-ratchet gear can be adapted to the special toothing with a first gear.

As regards terminology, the following should be noted. If any toothing drives another toothing, without load or performance, then this toothing pulls the driven toothing on the traction edge of the driving toothing on the traction edge of the driven toothing in the sense of the current drive or circumferential direction. If it comes to envelopes of the edge systems - such as when rattling or train-push-load change reactions - then come the trailing edges of these gears for engagement. This terminology changes when the direction of rotation or the drive or circumferential direction of this gear pairing changes.

In the context of the present application, the traction edge of the driving toothing is also referred to as the shear flank, and the traction flank of the driven toothing is also referred to as the trailing flank. Similarly, in the context of the present application, the non-engaged trailing edges are also referred to as a trailing edge or trailing edge.

Furthermore, the number of teeth and / or the tooth pitch (module) of the teeth and the respectively associated anti-rattle toothings are preferably identical. In contrast to anti-rattle measures in which, for example, the anti-rattle toothing of the anti-ratchet tooth has a tooth more or less than the associated toothing, constant tensions and a concomitant deterioration in efficiency are avoided due to the identical number of teeth or tooth pitch. (Optionally, however, the anti-rattle toothing may have the same pitch but fewer teeth than the associated toothing, with an anti-rattle tooth associated with only every second, third, fourth (generally nth) tooth of the tooth, particularly as a function of the frequency range of the spurious excitation; a sufficient jump coverage should be given).

It is also preferred if the toothings and the respectively associated anti-rattle toothings are substantially identical with respect to other toothing properties, for example with regard to the type of toothing (eg involute toothing), the helix angle, the tip circle diameter, the pitch circle diameter, the pressure angle, etc.

The above object is further achieved by a transmission with such a gear pair, in particular a motor vehicle transmission, and by a drive train with such a transmission.

If the motor vehicle transmission is designed as a stepped transmission in countershaft design, one or more sets of wheels (each with a loose wheel and at least one fixed gear) may have a gear pair according to the invention. The rattle tendency of such transmissions can be reduced to such an extent by the measures according to the invention that the multistage transmission can be designed without a dual mass flywheel (DMF). This leads to a significantly higher efficiency, since the total mass or the so-called mass factor can be reduced. The considerable costs for a ZMS can also be saved in this way. Furthermore, the tendency to spin up in the case of ZMS can be reduced. In addition, a better response can be achieved by reducing the rotational masses to be accelerated (the vehicle so equipped "better hangs on the gas"). A drive train with such a stepped transmission can have an adapted starting clutch, which has an integrated torsion damper (torsionally damped clutch disc). The multi-step transmission can be a manual, an automated or a dual-clutch transmission.

If the transmission is designed as a torque converter, at least one of the planetary gear sets may have a gear pairing according to the invention. The thus reduced rattle tendency can be used to more frequently close (earlier) a bypass clutch bridging the hydrodynamic converter. As a result, the efficiency can be increased.

The task is thus completely solved.

Preferably, the radial stop limits movement of the anti-ratchet gear with respect to the second gear such that maximum deflection of radially elastic deformation or bearing is established.

In this way, it can be avoided that the radially elastic deformation or mounting, as is established, for example, by a radial spring element, becomes too great. In other words, the radial stop is realized independently of the radially elastic deformation or storage.

It is advantageous if the second gear has a first, axially extending shoulder and the anti-ratchet gear has a second, axially extending shoulder, the shoulders form the radial stop.

As a result, the radial stop can be structurally set up comparatively easily. Preferably, the shoulders are formed in the region of mutually facing, radially aligned surfaces of the second gear or anti-ratchet gear.

So it is of particular advantage if the second gear has an axial recess which forms the first shoulder.

Such a recess can be comparatively easy to provide on a gear. In this case, the axial recess is preferably provided radially outside of an optionally provided annular projection, on which the anti-ratchet gear is supported in the radial direction.

Correspondingly, it is preferred if the anti-ratchet gear has an axial projection forming the second shoulder.

In the anti-ratchet wheel, which is generally relatively narrow in the axial direction, such an axial projection can be realized comparatively easily.

Alternatively, however, it is conceivable in the same way that an axial projection is provided on the second gear, whereas an axial recess is provided on the anti-ratchet gear.

Overall, it is preferable if the maximum eccentricity defined by the radial stop between the second gear and the anti-ratchet gear is smaller than the maximum eccentricity resulting from the engagement of the anti-rattle gears.

As a result, the maximum eccentricity established by the radial stop can be set to a value slightly greater than the eccentricity achieved during normal operation, which results, for example, from pushing the anti-ratchet wheel radially out of the first anti-rattle toothing in the region of meshing of the anti-rattle teeth. The maximum eccentricity defined by the radial stop can, for example, be less than 1 mm, in particular less than 0.5 mm, and in particular less than 0.25 mm. It is also particularly preferred if the maximum eccentricity is between 5-10 times greater than a difference between the teeth thickness second gear and the thickness of the teeth of the anti-rattle gear is.

According to a further preferred embodiment, the anti-rattle toothed wheel is arranged or designed such that a permanent two-flank rolling engagement is provided between the first and the second anti-rattle toothing.

In other words, the first and the second anti-rattle toothing are in rolling engagement such that, for example, always at least one tooth of the second anti-rattle toothing touches the two opposite flanks of a tooth gap of the first anti-rattle toothing. Although it is generally also conceivable that anti-rattle toothing interpret each other so that between them also a certain backlash prevails (as it is usually present in the toothing). In this case, however, the backlash of anti-rattle teeth is preferably smaller than the backlash of the teeth.

By the two-flank Wälzeingriff also relative movements of the components can be damped in the drive direction in both directions.

In general, the two-flank rolling engagement can be realized in any desired manner, for example by a positive profile displacement and / or the fact that the second anti-rattle toothing has a larger pitch circle diameter than the associated second toothing.

However, it is particularly preferred if the teeth of one of the anti-rattle toothings have a tooth thickness which is greater than or equal to the tooth gap of the teeth of the other anti-rattle toothing.

In other words, the two-flank rolling engagement is realized by making the backlash between the anti-rattle teeth zero and negative, respectively. If the tooth thickness is greater than the tooth gap, then the anti-ratchet gear is pushed away in the radial direction in the region of the meshing of anti-rattle toothing, from the first gear away and to the second gear out.

As mentioned above, a frictional force between the anti-ratchet gear and the second gear in the driving direction can thereby be increased.

According to a further preferred embodiment, the tooth thickness of the teeth of one of the anti-rattle toothings is greater by 20 μm to 500 μm, in particular by 50 μm to 250 μm, than the tooth thickness of the teeth of the associated toothed wheel and / or by a corresponding profile displacement.

This embodiment is particularly advantageous when the first anti-rattle toothing is formed by the first toothing. By this Zahndickenaufmaß can be achieved, for example, that the tooth thickness of the second anti-rattle toothing is greater than the largest tooth gap of the first anti-rattle toothing, under all operating conditions and all boundary conditions (functional, production-related (tolerances and any thermal and / or mechanical deformation of the components) ,

It is furthermore of particular advantage if the radial deflection of the anti-ratchet toothed wheel in the region of the meshing engagement with the first anti-rattle toothing is smaller than 500 μm, in particular smaller than 250 μm, particularly preferably smaller than 150 μm. The radial deflection of the anti-ratchet gear in the region of the tooth engagement is in particular smaller than the maximum eccentricity, which is established by the radial stop.

By dimensioning the anti-rattle teeth such that only such a small radial deflection is achieved, the efficiency of the anti-rattle measure can be very high, even if the hereby introduced into the anti-rattle gear forces are sufficient to reduce rattling or rattling of the teeth and preferably to prevent.

According to a further preferred embodiment, a spring rate with which the anti-rattle toothed wheel is biased in the radial direction or elastically deflectable in the radial direction is in the range of 2 to 100 N / mm, in particular in the range of 5 to 20 N / mm.

It has been found that such spring rates can reduce clamping effects and, on the other hand, the effectiveness of the anti-tampering measure can be high.

On the other hand, the desired damping characteristic for preventing rattling of the teeth can be surely achieved.

The above dimensions concerning the tooth thickness, the radial deflection and the spring rate relate to a conventional motor vehicle transmission for passenger cars, in particular to a center distance of the shafts carrying the components in the range of 60 mm to 90 mm and / or to a maximum transmissible via the transmission torque in the range of 150 Nm to 300 Nm. For smaller or larger gear pairings, these values must be adjusted accordingly.

It is particularly advantageous, if between the anti-ratchet gear and the second Gear a radial spring element is arranged, by means of which the anti-ratchet gear is elastically deflectable in the radial direction or biased toward the first gear.

In this way, the rotational backlash of the teeth can be compensated for 100%, since the anti-rattle gear can be pressed radially into the first anti-rattle toothing of the first gear, so that a tooth of the second anti-rattle toothing with the opposite edges of teeth of the first anti-rattle toothing is engaged.

The radial spring element may be formed as an annular wave spring element, but may also be designed as a "coil" spring or as a rubber spring.

The radial spring element can also fulfill the function of the frictional engagement between the anti-ratchet gear and the second gear, provided that the second aspect of the present invention is realized.

It is preferred if the radial spring element is produced as an annular spring made of an elastically deformable material such as spring steel.

When using spring steel for the radial spring element, the gear pairing over wide temperature ranges (especially at very low or very high temperatures as they are common in motor vehicles) reliably produce an anti-rattle effect.

Further, the spring ring may be circumferentially closed, but is preferably interrupted in the circumferential direction at one point, in particular to facilitate mounting in a radial groove of the second gear.

A steel radial spring element also has the advantage that the lubricant compatibility is better than, for example, O-rings. The same applies with regard to aging resistance.

According to a further embodiment, the anti-rattle toothed wheel has a radial spring section and is supported radially on the second toothed wheel, so that the anti-ratchet toothed wheel is elastically deflectable or biased toward the first toothed wheel itself in the radial direction.

The function of the radial spring section is essentially the same as that of the separate radial spring element. The two embodiments can also be combined with each other.

When forming the anti-ratchet gear with a radial spring portion, however, the number of parts of the gear pair can be reduced.

Although it is generally preferred to connect the anti-ratchet gear and the second gear via a frictional engagement in the drive direction (see below), the anti-ratchet gear may also be positively connected to the second gear in the drive direction.

In this embodiment, it may be advantageous if the anti-rattle gear even in the drive direction has a certain elasticity between the second anti-rattle toothing and the locating portion of the anti-rattle toothed wheel in order to damp vibration excitations in the drive direction.

According to a further preferred embodiment, an axially protruding annular projection is formed on the second gear, which faces the anti-ratchet gear.

The annular projection can be used for example as a guide means for guiding the anti-ratchet gear in the drive direction. However, the annular projection may also have other functions.

Furthermore, it is advantageous if the anti-ratchet gear is arranged laterally next to the second gear.

This simplifies the construction, wherein an arrangement laterally next to the second gear should also be understood to mean that the anti-ratchet gear is guided on an axially projecting annular projection of the second gear. However, it is particularly preferred in this case if the anti-rattle toothings are arranged laterally next to the toothings.

The anti-rattle gear can be made of any suitable material, such. As steel, be made.

However, it is generally advantageous if the anti-rattle gear is made of plastic.

As a plastic, for example, polyamide can be used, which has a high strength and rigidity and a very good chemical resistance. Furthermore, polyamide has a high resistance to wear and good sliding properties.

The mechanical properties can be adapted by fiber composites with glass or carbon fibers, in particular to reduce the water absorption.

Preferably, polyolefin-based additives are added to ensure high impact resistance.

Furthermore, the anti-rattle gear made of plastic can be produced inexpensively. In addition, it is possible to manufacture the anti-rattle gear made of plastic with relatively high precision, so that the backlash between the anti-rattle teeth can be made smaller than the backlash between the teeth.

Therefore, it is preferable that the anti-rattle gear is manufactured with a higher precision than the second gear.

According to a further preferred embodiment, the second anti-rattle toothing and / or the first anti-rattle toothing have teeth which are elastically deformable in the drive direction.

In this way, when the anti-rattle teeth have a backlash, noises that may occur according to the second aspect when turning over the anti-rattle teeth can be damped. However, the elastic deformability of the teeth in the drive direction can be useful even in a two-flank rolling engagement between the anti-rattle teeth according to the first aspect, and for the reasons mentioned above.

It is further preferred if the second anti-rattle toothing has teeth which are formed by the tooth head with radial slots.

In this way, the elastic deformability can be increased even with relatively stiff plastics (or other materials of anti-ratchet gear, such as steel).

The first anti-rattle toothing is formed according to a preferred embodiment of the first gear.

As a result, the number of parts can be reduced.

It is preferred if the first anti-rattle toothing is aligned with the first toothing, in particular in the axial direction. In other words, in this case teeth of the first toothing can be aligned with teeth of the first anti-rattle toothing.

Furthermore, it is advantageous if the first anti-rattle toothing is part of the first toothing.

In this embodiment, the first toothing is generally wider than the second toothing, wherein the axially projecting part of the first toothing forms the first anti-rattle toothing.

In this way, the first gear can be manufactured inexpensively.

According to an alternative embodiment, the first anti-rattle toothing is formed on a mating component, which is rigidly fixed to the first gear.

In this embodiment, the anti-rattle teeth can be ideally matched to one another based on the geometry and / or the material selection.

It is particularly advantageous if the counter component is arranged laterally next to the first gear.

The above object is achieved according to the second aspect of the invention by a gear pair of a first gear with a first toothing and a second gear with a second toothing, which is in engagement with the first toothing to transmit a driving force in a drive direction via the toothings to the first gear further associated with a first anti-rattle toothing, wherein on the second gear anti-ratchet gear is mounted, which is mounted relative to the second gear displaceable in the drive direction and a second anti-rattle toothing, which is in engagement with the first anti-rattle toothing, wherein the anti-ratchet gear is movably supported in the radial direction with respect to the second gear, wherein a radial stop limits relative movement of the anti-ratchet gear in the radial direction with respect to the second component such that maximum deflection is established, the anti-ratchet gear so elastically deformable formed and / or is elastically mounted in the radial direction that it is pressed away in the radial direction of the first gear in the region of the meshing with the first anti-rattle toothing, wherein the Raidalanschlag over the circumference of the second gear has a uniform radius, and wherein the second gear and the anti-ratchet gear each having a friction portion is assigned, which are indirectly or directly in frictional engagement with each other.

According to the second aspect, a two-flank rolling engagement between the anti-rattle gears is not required. However, it is preferred if a backlash between the anti-rattle teeth is smaller than a backlash between the associated teeth.

The gear pairings according to the second aspect of the invention can be combined with the features of the gear pairs of the first aspect of the invention, unless otherwise mentioned herein.

Characterized in that the second gear and the anti-ratchet gear are in frictional engagement with each other, the anti-rattle gear is taken by the second gear in the stationary state, without effect-reducing tension effects occur. In the case of a vibration excitation, starting from the steady state, the trailing edge (also referred to as leading edge or as working edge) of the driving gear from a thrust edge (also referred to as trailing edge or as non-working edge) of the driven gear solve and even to Turn over the opposite flank (due to the generally existing backlash between the teeth). The relative movement between the two gears is delayed or damped by means of the anti-ratchet gear or the friction engagement between the anti-ratchet gear and the second gear.

The backlash between the anti-rattle teeth is preferably less than the backlash between the teeth. If the backlash between the anti-rattle teeth is greater than zero, it can be achieved that, with a delay of the driving gear, the second anti-rattle toothing initially turns over (due to the smaller backlash). During the further course of the movement of the driving gear toward the counter flank, this movement is then delayed due to the frictional engagement between the second gear and the anti-ratchet gear. In this way it can be achieved that the second gear does not turn over during handling or at least with a lower relative speed to the first gear. Since the anti-rattle toothing here in normal operation, so for example under train or thrust, not constantly engaged or preferably are not braced against each other, by the anti-rattle measures according to the present invention, no secondary tonal noise generated such. B. howling.

However, the backlash between the anti-rattle teeth can be zero even in the second aspect of the invention, so that a two-flank rolling contact is achieved.

In this way, a desired friction effect can be well adjusted.

The anti-rattle mechanism of the gear pairing according to the invention is not a gear in mechanical engineering sense but a support mechanism for retaining or decelerating or damping the otherwise backlash in the rotational backlash and forth toothings. The anti-rattle measure is characterized by a high efficiency, since the mechanism acts only in the swinging (folding) of the teeth, but otherwise only internal forces between the anti-ratchet gear and the second gear act. Since the anti-rattle toothed wheel can be designed to be relatively narrow (for example in the range from 0.5 to 8 mm, in particular from 1 to 5 mm), there are also no significantly increased churning losses.

The anti-rattle mechanism can be provided with low weight and at low cost. Noise such as howling is not generated. When load changes (ie low-frequency turnover) the anti-rattle mechanism is overprinted. Since the mechanism in this case only during the phase of turning over the gears in effect, this results in no deterioration in the efficiency.

In contrast to prior art measures, in the second aspect of the invention, the anti-ratchet gear and the second gear are movable toward each other in the drive direction. Since the anti-rattle gear and the second gear on the friction grip in an operative relationship, thus the free flight phase of the teeth can be minimized when turning, especially when the backlash between the anti-rattle teeth is smaller than the backlash between the teeth.

The first and the second anti-rattle toothing come into positive engagement with each other according to the type of toothing. However, in the second aspect of the invention, the anti-rattle gears are generally not required to be involute gears such as the gears. Rather, the contour of the teeth of the anti-rattle teeth can be profiled spherical or convex. Ideally, the teeth of the anti-rattle teeth are in one point or in a line. Any profile pairings (convex-convex, plano-convex or convex-plan) are conceivable.

While in the second aspect of the present invention the anti-ratchet gear is preferably in frictional engagement with the second gear (ie frictionally engaged with the second gear in the stationary state in the drive state), in an alternative embodiment of the first aspect it is also possible in the invention to to connect the anti-ratchet gear in the drive direction positively to the second gear.

In this embodiment, the anti-rattle property of the anti-rattle gear can be realized substantially via the radial and / or tangential elasticity of the anti-rattle gear.

It is further understood that also in this embodiment, as well as in all other preceding embodiments, the anti-rattle toothing of anti-ratchet gear in the drive direction (ie tangential) may be formed elastically to dampen relative movements of the first and second component.

Overall, the following is to be noted: The aim is as 100% clearance as possible by the anti-rattle gear, preferably with small forces (about up to 50 N, in particular up to 30 N and more preferably up to 10 N) and small ways (especially smaller 20 microns, preferably about 2-4 microns) is pressed into the Gegenzahnradlücke. The rolling effects of circumferential gears are not hindered, tonal and other stochastic noise effects, especially rattles are avoided or at least significantly reduced.

The following aspects are therefore important:

• The application can be applied to any circumferential rolling external or internal gears, ie spur gear, planetary gear

• The application can also be used with any spline toothing, such as coupling splines / splines,

• The application can also be used with any dog teeth

• The application can also be used in general with any shaft-hub connections, for example as a supplement or replacement of the feather key. Their disadvantage is that it is always game-afflicted. This is disadvantageous in the case of oscillation of one or the other component of the shaft-hub connection, since due to the play / nonlinearities, impact effects / rattling can occur. So if, for example, instead of the classical feather key after the type of anti-ratchet (or micro gear), such as a tooth as extreme simplification of the micro gear, radially and / or tangentially elastic pressed into the mating tooth gap or keyway of the Gegenbauteiles (shaft or hub), it can not Rattle effects can occur or they can be alleviated.

The forces of the micro gear can preferably be generated so that the anti-ratchet gear (micro gear) is pressed with two flank rolling engagement due to its thicker teeth opposite the largest mating gear gap radially elastic in the mating gear gap. The forces applied for this purpose can generally be generated arbitrarily by magnetism, spring force, hydraulics, pneumatics, etc., even if only the variant of spring elasticity is presented here.

The arrangement of the Antirasselzahnrads (micro gear) can be done at Geradverzahnungen arbitrarily on one or the other or both sides of the mother gear simultaneously. In helical gears can be depending on the direction of the helix angle preferential sides of the axial arrangement, such as one or the other side of the nut gear. Because in this case, the micro gear could be either pressed axially on the nut gear or pushed away from this as a result of the resulting superimposition of forces. Overall, however, the executed constructions are preferably to be dimensioned so that, no matter what forces act, the effect of rattling is suppressed as 100% as possible, the radial and / or Tangentialscherelastizität is given and / or the micro gear non-positively or positively in position remains unintentionally pressed axially against the nut gear or pushed away from it. In some designs made it may be that there is no free choice of the arrangement of the micro gear on any face of the nut gear. If an unfavorable position of the microgear thus results in such constraints, and thus also unfavorable superpositions of forces, then suitable constructions are to be selected, such as support rims, spring washers, etc., so that the microgear is not unintentionally pushed axially away from the nut gear.

The anti-ratchet gear is preferably designed as a ring element. Preferably, the ratio of outer diameter to inner diameter of the ring element is in the range of 100: 50 to 100: 95, in particular in the range of 100: 60 to 100: 85, in particular in the range of 100: 70 to 100: 80. As a result, the anti-rattle gear can be formed with a low weight.

It is understood that the features mentioned above and those yet to be explained below can be used not only in the particular combination given, but also in other combinations or in isolation, without departing from the scope of the present invention.

Embodiments of the invention are illustrated in the drawings and are explained in more detail in the following description. Show it:

1 a schematic longitudinal sectional view through an embodiment of a gear pair according to the invention;

2 a partial sectional view corresponding to 1 by a further embodiment of a gear pair according to the invention;

3 a longitudinal sectional view through a further embodiment of a gear pair according to the invention;

4 a schematic development of a gear pair of a fixed gear and a loose wheel according to a first embodiment of the invention;

5 one of the 4 comparable view, with a trailing edge of the fixed wheel detaches from a trailing edge of the idler gear;

6 one of the 4 comparable view, wherein a trailing edge of the fixed wheel rests against a trailing edge of the idler gear;

7 a schematic representation of a gear pair according to the invention;

8th a schematic representation of a tooth engagement of a gear pair according to the invention;

9 a schematic representation of a first drive train for a motor vehicle; and

10 a schematic representation of another drive train for a motor vehicle.

A first embodiment of a mechanical gear pairing according to the invention is shown in FIG 1 represented and generally with 10 designated.

The gear pairing 10 has a first gear 12 in the form of a loose wheel (rotatable about an axis 13 ) and a second gear 14 in the form of a fixed wheel. The fixed wheel 14 is in this case the driving gear and on a shaft 15 established. The idler wheel 12 has a first toothing 16 on. The fixed wheel 14 has a second toothing 18 on.

To prevent rattle noise is the gear pairing 10 equipped with an anti-rattle mechanism, which is an anti-rattle gear 30 includes. The anti-rattle gear 30 is with the fixed wheel 14 coupled, in such a way that the anti-rattle gear 30 in the drive direction 20 movable against the fixed wheel 14 is. The anti-rattle gear 30 stands with a first anti-rattle toothing 32 engaged on the idler gear 12 is provided. For this purpose, the anti-rattle gear has 30 a second anti-rattle toothing 33 on.

The first anti-rattle toothing 32 can be an axial section of the first toothing 16 be. The first anti-rattle toothing 32 However, it can also be shaped differently than the first toothing 16 but axially aligned therewith.

The anti-rattle gear 30 is manufactured with a relatively high precision, such that a circumferential play 34 between the anti-rattle teeth 32 . 33 is smaller than the backlash 24 , and preferably less than or equal to zero, so that a two-flank rolling engagement is realized.

The anti-rattle gear 30 is on the second gear 14 via a radial spring element 40 in the form of a spring (in the present case a corrugated spring) mounted radially elastic. The second gear (fixed wheel) 14 has a ring projection 42 on, on which the radial spring element 40 supported in the radial direction. The radial spring element 40 is on the inner circumference of the anti-rattle gear 30 in a circumferential groove 44 added. Further, on the outer periphery of the annular projection 42 a radial groove 46 provided, in which the inner circumference of the radial spring element 40 attacks. Through the circumferential groove 44 is the radial spring element 40 in the axial direction positively with the anti-ratchet gear 30 coupled. Through the radial groove 46 is the radial spring element 40 in the axial direction positively with the fixed gear 14 coupled.

The shape of the grooves 44 . 46 may be arbitrary, but it is preferably rectangular in cross-section grooves. Instead of a wave spring can also any other radial spring element 40 be provided by means of which the anti-rattle gear 30 on the second gear 14 is mounted radially elastic.

The second anti-rattle toothing 33 is preferably formed so that it has a two-flank rolling contact with the first anti-rattle toothing 32 forms. Preferably, this is between the anti-rattle teeth 32 . 33 equipped backlash smaller than zero, so the anti-rattle gear 30 in the region of the tooth engagement in a radial direction 115 is pushed away from the tooth engagement, against the radially elastic effect of the radial spring element 40 , In other words, the anti-rattle gear 30 in the radial direction against the anti-rattle toothing 32 of the first gear 12 be biased.

Preferably, between the radial spring element 40 and the anti-rattle gear 30 a frictional engagement 38a be realized, in particular in the area of the circumferential groove 44 , In a similar way, a frictional engagement 38b between the Radial spring element 40 and the fixed wheel 14 be set up, in the radial groove 46 , This is the anti-rattle gear 30 in the circumferential direction 20 relative to the fixed wheel 14 agile, through the frictional interventions 38a . 38b ,

Alternatively, the radial spring element 40 but in the circumferential direction 20 also form-fit with the anti-ratchet gear 30 and / or with the fixed wheel 14 be coupled. In this case, the cross section of the grooves 44 . 46 be formed, for example, corrugated. In this case, a rattling alone by the two-flank rolling engagement between the anti-rattle teeth 32 . 33 be reduced or eliminated.

To limit the relative radial movement of the anti-rattle gear 30 in relation to the second gear 14 is also a radial stop 50 intended. The radial stop 50 is independent of the radial spring element 40 provided and is adapted to the radially elastic deformation of the radial spring element 40 to limit to a maximum deflection. The radial stop 50 is formed in the region of facing end faces of the anti-ratchet gear 30 and the second gear 14 , More specifically, the second gear 14 an axial recess 56 on, the radial outer diameter of a first shoulder 52 forms. The anti-rattle gear 30 has an axial projection 58 on, in the axial recess 56 engages and on its radial outer circumference a second shoulder 54 forms. With concentric alignment of the second gear 14 and the anti-rattle gear 30 (as in 1 shown) are the shoulders 52 . 54 spaced from each other to the radial mobility of the anti-ratchet gear 30 to ensure. The maximum radial deflection or maximum eccentricity is in 1 With 60 designated and then set up when the second shoulder 54 the first shoulder 52 touched.

Because of the pushing out of the anti-ratchet gear 30 from the first anti-rattle toothing 32 the anti-rattle gear 30 in relation to the second gear 14 circulates eccentrically, so to speak (see also 7 ), centrifugal or centripetal forces inevitably occur. Through the radial stop 50 In this case, it is reliably prevented that these forces under all possible operating conditions can not lead to the anti-ratchet wheel 30 this prescribed circular path or wobble path with respect to the second gear 14 leaves. Through the radial stop 50 Consequently, the reliability can be increased.

The radial stop 50 is therefore particularly advantageous when the anti-rattle gear 30 by means of the radial spring element 40 in the radial direction against the first anti-rattle toothing 32 is biased. The radial stop 50 However, it may also be advantageous if between the anti-ratchet gear 30 and the second gear 14 Although no radial bias is established, but a radial mobility is possible. In these cases too, centrifugal or centripetal forces may occur in certain operating situations, the effects of which due to the radial stop 50 be limited.

While the anti-rattle gear 30 in the embodiment of the 1 over the radial spring element 40 axially on the second gear 14 is fixed, such axial fixation can also be done by other means (for example via an axial securing ring or the like.). Furthermore, in 1 shown a frictional engagement between the anti-rattle gear 30 and the second gear 14 can also be set up in the region of the abutting end faces. Such a frictional engagement 38 can also by an axial pressure force 36 be set up or set.

Although it is preferred for structural reasons, if an axial recess 56 on the second gear 14 and an axial projection 58 on the anti-rattle gear 30 is provided, in an appropriate manner, an axial recess on the anti-ratchet gear and an axial projection on the second gear 14 be provided to the radial stop 50 to set up.

In 2 is shown in schematic form such an embodiment in which a radial stop 50 is formed by a radially inner shoulder 52 a projection of the second gear 14 and by a corresponding radially inner shoulder 54 an axial recess in the anti-rattle gear 30 ,

In 2 is further shown that a second radial stop between a radially outer shoulder 62 the projection of the second gear 14 and a corresponding shoulder 64 on the outer circumference of the recess of the anti-ratchet gear 30 can be set up. By this measure, the radial mobility in the direction radially outward and the radial mobility in the direction radially inward can be limited independently.

In the 1 shown gear pairing 10 has two gears 12 . 14 on, which are straight toothed. However, the gear pairing according to the invention can also be designed as a pairing of a rack and a gear.

In 3 is another embodiment of a gear pair 10 shown in general in terms of structure and operation of the gear pair of 1 equivalent. The same elements are therefore identified by the same reference numerals. The following section essentially explains the differences.

In 3 is on the inner circumference of the anti-rattle gear 30 no circumferential groove 44 provided, the axial positive engagement between the radial spring element 40 and the anti-rattle gear 30 sets up. Rather, the radial spring element 40 designed so that it attaches to the second gear 14 facing away from the anti-rattle gear 30 abuts and due to its support in the circumferential groove 46 of the annular projection 42 an axial pressure force 36 on the anti-rattle gear 30 exerts, so a frictional engagement 38 between the mutually facing end faces of the anti-ratchet gear 30 and the second gear 14 is set up. Here is the radial groove 46 Semicircular in longitudinal section, and the radial spring element 40 has on its inner circumference a corresponding semicircular shape, via which a positive connection in the axial direction between the second gear 14 and the radial spring element 40 is set up. The radial spring element 40 consequently also serves for the axial securing of the anti-ratchet gear 30 ,

In the 4 to 6 is an embodiment of a gear pair 10 shown at the time between the first anti-rattle toothing 32 and the second anti-rattle toothing 33 no two-way rolling contact is set up. Rather, between these gears is a circumferential game 34 However, this is smaller than a backlash between the teeth 16 . 18 ,

The representations of the 4 to 6 serve to explain the rattle phenomenon and to explain how the rattle phenomenon according to a second aspect substantially due to a frictional engagement 38 between the anti-rattle gear 30 and the second gear 14 can be reduced. However, the illustration is also in a corresponding manner to the embodiment of 1 to 3 applicable, in which a two-flank rolling contact is established, in which the circumferential clearance 34 Zero or less than zero.

The idler wheel 12 is by means of the fixed wheel 14 in a drive direction 20 driven. It touches a shear flank 22 the second gearing 18 a trailing edge 26 the first gearing 16 ,

The gears 16 . 18 are designed with a certain backlash, which in 4 With 24 is designated.

The backlash 24 is in the present case, the distance between a trailing edge of the second toothing 18 and a trailing edge 28 the first gearing 16 ,

Such gears are well known. Due to the backlash 24 In the case of higher-frequency excitations on the drive side, so-called rattle noise can occur. This is where the gears beat 16 . 18 around, so that alternately the flanks 27 . 28 and the flanks 22 . 26 touch.

In particular, such high-frequency excitations when using such a gear pairing 10 occur in a drive train of a motor vehicle, for example in a stepped or helical gear of such a drive train. This is especially true when the transmission is coupled on the input side with a drive motor that generates vibrations, such as an internal combustion engine.

The anti-rattle gear 30 is also on the fixed wheel 14 over the frictional engagement 38 guided. These are on the anti-rattle gear 30 (or related components) and on the fixed wheel 14 (or related components) corresponding friction surfaces formed, preferably by an axial pressure force 36 be engaged with each other. The corresponding representation is in the 4 to 6 schematic nature and is merely intended to indicate that the anti-rattle gear 30 in the drive direction 20 relative to the fixed wheel 14 can be moved, but in this case a certain frictional force due to the frictional engagement 38 to overcome. This anti-rattle mechanism can significantly reduce or even completely eliminate rattle noise as described above.

If, due to a higher-frequency excitation, the trailing edge 22 the second gearing 18 from the trailing edge 26 the first gearing 16 Rist, due to the frictional engagement becomes the anti-rattle gear 30 taken here. At some point in time 2 is shown beats the trailing edge 26 the second anti-rattle toothing 33 at a corresponding shear flank 22 the first anti-rattle toothing 32 on (the big game 34 is overcome).

Due to the higher-frequency excitation becomes the fixed wheel 14 then moved further in the direction of the envelope. However, this movement is due to the frictional engagement 38 delayed or braked or damped. As a result, the trailing edge hits 27 the fixed wheel 14 at a significantly reduced speed (ideally zero speed or not at all) on the trailing edge 28 of the idler wheel 16 on.

In this way, rattle noise, as they occur in conventional gear pairings, can be reduced efficiently.

The anti-rattle gear 30 is preferably made of plastic, in particular of polyamide. The first gear 12 and the second gear 14 are preferably made of metal, for example of steel alloys using chromium, nickel, molybdenum, etc.

During the return movement of the fixed wheel 14 in terms of idler gear 12 the same procedure takes place. The anti-rattle gear 30 is first of the Festrad 14 taken along, until its flank 18 on the trailing edge 26 the first anti-rattle toothing 32 strikes. As a result, the further movement of the fixed wheel 14 on the loose wheel 12 in turn due to the frictional engagement 38 delayed. Therefore, it can be achieved that the shear flank 22 then with a low speed on the trailing edge 26 impinges (or in the ideal case with the speed zero or not at all).

In 7 is in schematic form a gear pairing 10 shown at the anti-rattle gear 30 in the radial direction 115 is elastically deflectable. This is the anti-rattle gear 30 overall rigid and radially elastic mounted. For example, due to a larger tooth thickness (see below) becomes the anti-rattle gear 30 doing so in the radial direction 115 from the anti-rattle toothing 32 of the first gear pushed out. This results in a radial offset between the teeth in the region of the meshing 150 which is generally smaller than the maximum eccentricity 60 (which may be, for example, 1 mm).

Because the anti-rattle gear 30 is formed substantially rigid, this is with respect to the second gear 40 offset eccentrically so that their centers are radially offset, as in 152 is shown.

In an alternative embodiment, the anti-rattle gear 30 For example, be self-elastic. This results in turn in the region of the meshing a radial offset 150 whereas, on the radially opposite side, such radial misalignment does not have to exist.

In 8th is in schematic form a preferred embodiment of a gear pair 10 shown. Here, for example, the second anti-rattle gear 14 with a driving force 160 driven in the drive direction. In this case, there is a flank of the second toothing 18 on a flank of the first toothing 16 at. There will be a driving force 162 on the first gear 12 transfer. This takes place in one place 164 the meshing between the teeth 16 . 18 instead of. In 8th is also the backlash 24 between the gears 16 . 18 shown.

The anti-rattle toothing 33 of anti-rattle gear 30 On the other hand, it is designed so that it fits with the first anti-rattle toothing 32 of the first gear 12 is in a two-flank rolling engagement. Here find a meshing between these teeth on the one hand in one place 166 instead, for example, with the place 164 can coincide. On the other hand, the teeth touch 33 . 32 also on an opposite flank, what with 168 is shown.

The teeth of the second anti-rattle toothing 32 are designed so that they have a tooth thickness 170 which is larger than a tooth gap 172 the first anti-rattle toothing 32 , This causes the tooth over the teeth 166 . 168 in the radial direction 115 from the first anti-rattle toothing 32 is pressed out, against the force of a radial spring element shown schematically. The consequent radial deflection is in 38 again with 150 shown. In relation to 8th It should be noted that the difference between the tooth thickness 170 and the tooth gap 172 is exaggerated enlarged to illustrate the facts more clearly. Consequently, the radial offset is also 150 already exaggerated. As a rule, this is smaller than 500 μm.

The anti-rattle teeth 32 . 33 are designed as involute gears. The information of the tooth thickness and the tooth gap refer to the usual nomenclature on the tooth thickness in the so-called pitch circle.

The above-described embodiments of gear pairings fulfill at least one of the following advantages:
The problem of rattle on a gear pair with play is solved by only one gear in one plane.

A clearance compensation is done either by a profile shift, by thicker teeth of anti-rattle toothing 33 of anti-rattle gear 30 , by a smaller tooth gap of the anti-rattle toothing 32 , by radially impressing the anti-ratchet gear 30 in the anti-rattle toothing 32 of the first gear, by increasing the volume of the anti-rattle gear until tight meshing is achieved, by resilient springback action into the anti-ratchet gear 30 are integrated, for example in the radial and / or tangential (ie in the drive direction) direction, by frictional engagement of the anti-ratchet gear 30 in the axial or radial direction relative to the second gear 14 . wherein the frictional engagement can take place directly or indirectly, and / or by an axial or radial fixation of the anti-rattle toothed wheel 30 on the second gear 14 ,

The anti-rattle gear 30 may be associated with a complementary anti-ratchet gear, either by positive or non-positive connection. For example, the complementary anti-ratchet gear may be of a different material than the anti-rattle gear. For example, the complementary anti-rattle gear 90 be made of plastic (eg., Polyimide), whereas the anti-rattle gear 30 even made of metal. This can minimize the amount of metal (steel) use to reduce cost and weight.

When using metal for the anti-rattle gear, the use of aluminum (or another light metal) instead of steel is conceivable. In one embodiment, the anti-rattle gear is 30 preferably formed of plastic or a light metal, as a result, a desired thermal expansion compensation is possible.

In a gearbox with a plurality of sets of wheels, an anti-rattle measure is preferably formed on at least one, preferably each of the wheel sets, as above. In this case, it is preferable if the drive train in which the transmission is used has no dual-mass flywheel on the output side of the internal combustion engine. For this case, it is preferred if not only the wheelsets are formed with a anti-rattle measure, as described above, but also if rattle vibrations of the synchronizer rings are reduced in the transmission, for example by clipping corrugated springs between the coupling body and synchronizer ring.

In the 9 and 10 exemplary drive trains for motor vehicles are shown in which the gear pairing according to the invention can be used.

9 shows in schematic form a drive train 180 for a motor vehicle that has an internal combustion engine 182 and a starting clutch 184 having. Furthermore, the powertrain includes 180 an executed in Vorgelegebauweise step transmission 186 in the usual way a plurality of wheelsets 188 includes. The wheelsets 188 are switchable by means of clutches (synchronous clutches) to different gears of the stepped transmission 186 to design or interpret. The wheelsets 188 typically include a constant gear set and a plurality of sets of wheels each including a loose wheel and one or more fixed wheels.

In 9 is further exemplified that at least one of the wheelsets 188 a gear pairing 10 according to the present invention. The wheelset 188 includes a first gear 12 in the form of a loose wheel, which is switchable by means of a synchronizer clutch, and a second gear 14 in the form of a fixed wheel. The fixed wheel 14 is an anti-rattle gear 30 associated with the type according to the invention.

10 shows an alternative embodiment of a drive train 200 for a motor vehicle having a drive motor 182 , a hydrodynamic converter 202 and a planetary gear 204 includes. The planetary gear 204 includes at least one planetary gear set 206 , which is switchable by unspecified clutches or brakes. In this case, for example, form the planetary gears of the planetary gear 206 second components 14 in the sense of a gear pair according to the invention. The sun gear is the first gear 12A formed, the ring gear is also a first gear 12B educated. The planetary gears (the second components) 14 stand with both the sun gear 12A as well as with the ring gear 12B engaged. In this case, at least one of the planet gears 14 an anti-rattle gear 30 be assigned according to the present invention.

In the powertrain 180 of the 9 It is advantageous that the drive train between the drive motor 112 and the clutch 184 does not have to include a dual mass flywheel. However, the clutch can 184 itself be designed with a conventional torsional damper, which may include a two- or multi-stage characteristic.

In the powertrain 200 It is advantageous that a lock-up clutch 208 for bridging the hydrodynamic converter 202 can be switched on more frequently or earlier, so that the efficiency of the drive train 200 can be increased. In addition to the application in gearboxes of gearboxes, the following applications are generally conceivable: motor control wheels, industrial gear units, pumps, gear pumps, machine tools, household appliances, Lifescience products such as electric toothbrushes, kitchen machines. The use in transmissions is not limited to the use in passenger cars, but also tunable for use in transmissions for commercial vehicles.

With regard to the dimensioning of the gear pairs according to the invention, the following should also be noted. In each application, the dimensions or geometries according to the required physical principles of action are each individually determined by usable calculation approaches and - if they are not sufficiently known or existed - by empirical experiment votes exactly so that the required function of the function carrier / components in each conceivable functional case full and as desired - as described above - is met.

The named numerical values apply in particular to a manual transmission of a passenger vehicle with a displacement of 1.6 liters and a maximum transmittable torque of 217 Nm. The main axle is between drive shaft and secondary shaft 72 mm. Any other interpretation of the gear pairing must be individually re-tuned. It is approximately true that the parameters speed, amplitude of the angular acceleration and moments of inertia in a rational ratio linearly affect these named forces and spring stiffness of the gear pair. To put it in simple terms: double speed, or double the amplitude of the angular acceleration or double the mass moment of inertia of the loose part to be prevented by the gear pairing on rattling and rattling must be achieved with a doubling of the spring forces and stiffnesses. Conversely, the same applies to halving the named parameters. For the tooth thickness widening of the micro gear (at least in the first aspect of the invention) is generally independent of the size and design of the transmission: The tooth thickness of the micro gear must always be greater than ever occurring by manufacturing variations, thermal expansion or mechanical deformation tooth gap of the counter wheel, so always it is certainly ensured that the two-flank gear pairing of the micro gear eliminates any backlash to the mating gear by 100%.

Claims (15)

Gear pairing ( 10 ) with a first gear ( 12 ) with a first toothing ( 16 ) and a second gear ( 14 ) with a second toothing ( 18 ), with the first gearing ( 16 ) is engaged to engage over the gears ( 16 . 18 ) a driving force in one direction of rotation ( 20 ), wherein the first gear ( 12 ) furthermore a first anti-rattle toothing ( 32 ), wherein on the second gear ( 14 ) an anti-rattle gear ( 30 ), which is a second anti-rattle toothing ( 33 ), which with the first anti-rattle toothing ( 32 ), wherein the anti-rattle gear ( 30 ) in the radial direction ( 115 ) movable with respect to the second gear ( 14 ), wherein a radial stop ( 50 ) a relative movement of the anti-rattle gear ( 30 ) in the radial direction with respect to the second gear ( 14 ) so limited that a maximum deflection ( 60 ), characterized in that the anti-rattle gear ( 30 ) is formed so elastically deformable and / or is elastically mounted in the radial direction, that in the region of the tooth engagement with the first anti-rattle toothing ( 32 ) in the radial direction ( 115 ) of the first gear ( 12 ) is pushed away, wherein the radial stop ( 50 ) over the circumference of the second gear ( 14 ) has a uniform radius. Gear pairing according to claim 1, characterized in that the number of teeth of the anti-ratchet gear equal to the number of teeth of the second gear ( 14 ). Gear pair according to claim 1 or 2, characterized in that the second gear ( 14 ) a first, axially extending shoulder ( 52 ) and the anti-rattle gear has a second, axially extending shoulder ( 54 ), the shoulders ( 52 . 54 ) the radial stop ( 50 ) form. Gear pair according to claim 3, characterized in that the second gear ( 14 ) an axial recess ( 56 ), which the first shoulder ( 52 ). Gear pair according to claim 3 or 4, characterized in that the anti-rattle gear ( 30 ) an axial projection ( 58 ), the second shoulder ( 54 ). Gear pair according to one of claims 1-5, characterized in that by the radial stop ( 50 ) defined maximum eccentricity ( 60 ) between the second gear ( 14 ) and the anti-rattle gear ( 30 ) is greater than the maximum eccentricity resulting from engagement of the anti-rattle teeth. Component pairing according to one of claims 1-6, wherein the anti-rattle gear ( 30 ) is arranged or formed such that between the first and the second anti-rattle toothing ( 32 . 33 ) a permanent two-flank rolling engagement is given. Component pairing according to any one of claims 1-7, wherein the teeth of a ( 33 ) of anti-rattle teeth ( 32 . 33 ) have a tooth thickness which is greater than or equal to the tooth gap of the teeth of the other anti-rattle toothing ( 32 ). A gear pair according to any one of claims 1-8, wherein the radial deflection of the anti-rattle gear ( 30 ) in the region of the tooth engagement with the first anti-rattle toothing is smaller than 500 μm, in particular smaller than 250 μm. A gear pair according to any one of claims 1-9, wherein between the anti-rattle gear ( 30 ) and the second gear ( 14 ) a radial spring element ( 40 ) is arranged, by means of which the Anti-rattle gear ( 30 ) in the radial direction ( 115 ) is elastically deflectable. A gear pair according to any one of claims 1-10, wherein the anti-rattle gear ( 30 ) in the drive direction relative to the second gear ( 14 ) is mounted displaceably. Gear pair according to one of claims 1-11, wherein the second gear ( 14 ) and the anti-rattle gear ( 30 ) is assigned in each case a friction section, which in frictional engagement ( 38 ) stand together. A gear pair according to any of claims 1-12, wherein on the second gear ( 14 ) an axially protruding annular projection ( 42 ) formed the anti-rattle gear ( 30 ) is facing. Transmission ( 186 ; 204 ) with at least one gear pair ( 10 ) according to any one of claims 1-13. Powertrain ( 180 ; 200 ) for a motor vehicle with a transmission ( 186 ; 204 ) according to claim 14.

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