WO1999039115A1

WO1999039115A1 - Infinitely adjustable rolling-contact gears

Info

Publication number: WO1999039115A1
Application number: PCT/EP1999/000625
Authority: WO
Inventors: Georg Albersinger
Original assignee: Georg Albersinger
Priority date: 1998-02-02
Filing date: 1999-02-01
Publication date: 1999-08-05
Also published as: DE19804011A1

Abstract

The invention relates to infinitely adjustable rolling-contact gears, preferably for use in vehicle manufacture, comprising at least one tapered disk (10, 11; 100, 110), a group of rotationally mounted planetary rollers (3; 30) which have at least one conical running surface (3a; 30a), and at least one group of double-cone-shaped, rotationally mounted intermediate rollers (5) which can be displaced longitudinally and are positioned between the planetary rollers (3; 30) and the tapered disk (10, 11; 100, 110). The tapered disks, planetary rollers and intermediate rollers are arranged in a rotationally symmetrical manner in relation to the longitudinal gear axis (1a). The running surfaces (10a, 11a; 110a; 3a; 30a) of the tapered disk (10, 11; 100, 110) and planetary rollers (3; 30) facing each other are embodied such that the cutting plane lines (FL1, FL2) of said running surfaces extending through the longitudinal gear axis (1a) in cutting planes are parallel to each other. The rotational axes (6a) of the intermediate rollers (5) are positioned at an angle to the cutting plane lines (FL1, FL2) of corresponding conical running surfaces (10a, 11a; 100a, 110a; 3a; 30a) of the tapered disk (10, 11; 100, 110) and planetary rollers (3, 30). The rotational axes (6a) of all intermediate rollers (5) of each group of intermediate rollers can be displaced jointly in relation to the tapered disk (10, 11; 100, 110) and the planetary rollers (3, 30) parallel to the longitudinal gear axis (1a), resulting in a change in the transmission ratio of the rolling-contact gears (1).

Description

CONTINUOUSLY ADJUSTABLE ROLL GEARBOX

DESCRIPTION

The invention relates to a continuously variable rolling gear.

A large number of solutions are known in the field of continuously variable drives for vehicles and machines. In the case of transmissions based on the frictional force principle, a known embodiment works according to the wrap-around principle, in which flexible, band-shaped transmission elements, for example steel chains or push link belts, are stretched between two pairs of disks, each consisting of two conical disks located opposite one another. By changing the distance between the conical disks of each pair of disks in opposite directions, the effective radii and thus the gear ratio can be changed continuously. Another known embodiment consists of a rolling gear based on toroids with two opposing profile disks serving for driving and driving. The profile disks each have a semi-annular recess on their opposite surfaces in order to delimit an annular or toroidal space in which a plurality of disk-shaped transmission members are mounted. Each disk-shaped transmission member touches the semi-ring-shaped bores of both profile disks (drive and driven disk) with its spherical peripheral surface. By pivoting each transmission member around its center, different effective radii with respect to the semi-ring-shaped bores of the drive and driven disks and thus different transmission ratios between the drive and driven disks can be set. The disadvantages of friction power transmissions based on the wrap principle consist primarily in the high operating noise, the relatively short service life and the relatively low adjustment speed combined with high contact forces for each conical disk pair. For these reasons, such transmissions have so far only been built for relatively low outputs.

In toroidal-based roller companies, there is a high proportion of drilling friction with linear contact between the disk-shaped transmission members and the walls of the semi-ring-shaped bores of both profile disks. Furthermore, because of the unfavorable spatial arrangement in the toroidal space, a synchronous adjustment of the disk-shaped transmission members is difficult to achieve technically.

In contrast, the object of the invention is to provide a continuously variable rolling gear which has the lowest possible drilling friction with the longest possible and many line-like frictional force contacts. Such a roller gear should preferably have a high power density so that it can be used in vehicle construction.

This object is achieved by the characterizing features of claim 1.

Advantageous refinements and developments of the continuously variable roller transmission according to the invention result from the subclaims. The invention is based on the consideration that low-friction transmission with adjustable transmission ratio can be achieved by tapered rolling elements. In the case of tapered rolling elements, all the important conditions for optimal operation of friction gears, namely long lines of friction with low drilling friction and variable translation, are combined. These basic principles can be implemented in various embodiments of the invention, which are shown in more detail in the drawings. It shows:

1 shows a perspective view of a first embodiment of the roller transmission according to the invention with external planet rollers, the axes of rotation of which are oriented essentially parallel to the longitudinal axis of the transmission;

FIG. 2 shows an axial section through the first embodiment of the roller transmission according to the invention according to FIG. 1;

FIG. 3 shows a perspective view of a second embodiment of a continuously variable roller transmission according to the invention, which is essentially formed by mirror-symmetrical doubling of the first embodiment according to FIG. 1;

FIG. 4 shows an axial section through the second embodiment of the roller transmission according to the invention according to FIG. 3;

5 and 6 are schematic views of the two end positions of the roller transmission according to FIG. 3; 4

FIG. 7 shows a schematic view of the roller transmission according to FIG. 3, in which the axis of rotation of the external planet rollers can be deflected to a small extent with respect to the longitudinal axis of the transmission in order to facilitate the change in the transmission ratio;

FIG. 8 shows a perspective view of a third embodiment of the roller transmission according to the invention with internal planet rollers, the axes of rotation of which are oriented essentially perpendicular to the longitudinal axis of the transmission;

Figure 9 is a perspective view of the centrally located

Supporting elements of the rolling gear according to Figure 8 for the planetary rollers and the intermediate rollers;

FIG. 10 shows an axial section through the third embodiment of the roller transmission according to the invention according to FIG. 8;

Fig. 11 u. 12 schematic views of the two end positions of the roller transmission according to FIG. 8;

FIG. 13 shows an axial view of a support element for the star-shaped planet rollers of the roller transmission according to FIG. 8, in which the frictional force between the individual tapered rollers of the roller transmission can be adjusted depending on the torque;

FIG. 14 shows a perspective illustration of a fourth embodiment of the roller transmission according to the invention, which is essentially based on the use of only one half of the roller transmission according to FIG. 8; FIG. 15 shows an axial section through the fourth embodiment of the roller transmission according to the invention according to FIG. 14;

FIG. 16 shows a perspective view of a fifth embodiment of the roller transmission according to the invention with internal planet rollers which, in a modification of the embodiment according to FIG. 8, have curved conical surfaces;

FIG. 17 shows an axial section through the fifth embodiment of the roller transmission according to the invention according to FIG. 16,

18 shows a perspective view of a sixth embodiment of the roller transmission according to the invention, which, in a modification of the embodiment according to FIG. 3, has concave-tapered running surfaces for the conical disks and planet rollers, and

FIG. 19 shows an axial section through the sixth embodiment of the roller transmission according to the invention according to FIG. 18.

The first embodiment of a steplessly adjustable roller transmission 1 shown in perspective in FIG. 1 and in section in FIG. 2 comprises two annular planet carriers 2a and 2b which are arranged rotationally symmetrically with respect to the longitudinal axis la of the roller transmission 1 at a mutual distance. In each planet carrier 2a, 2b, a group of bores are made uniformly distributed over the circumference, the axes of these bores running parallel to the longitudinal axis la of the rolling gear 1. The bores in the planet carriers 2a, 2b serve to receive support axes 4, which with their one end section in the planet carrier 2a are rotatably and axially immovable and are axially displaceably mounted with their other end section in the planet carrier 2b. The support axes 4 serve to receive a group of planet rollers 3, which are formed with a tapered running surface 3a. The smaller diameter of each planetary roller 3 is on the side of the planet carrier 2b. Each planetary roller 3 is mounted on its supporting axis 4 so as to be rotatable about an axis of rotation 4a parallel to the longitudinal axis la of the transmission and axially immovable. Furthermore, a spur gear 3b is rotatably mounted on each support axis 4 between the planetary roller 3 located there and the planet carrier 2a and is connected in a rotationally fixed manner to the planetary roller 3. All spur gears 3b mesh with a central ring gear 8a which is fastened on a central output shaft 8. The axis of rotation of the output shaft 8 coincides with the longitudinal axis of the transmission la.

At a fixed axial distance from the output shaft 8 is a central drive shaft 7, the axis of rotation of which also coincides with the longitudinal axis la of the transmission. The central drive shaft 7 carries at its end facing the output shaft 8 a conical pulley 11, the smallest diameter of which lies opposite the axial end of the output shaft 8. It is essential that the cone angle of the cone pulley 11 corresponds exactly to the cone angle that is the same for all planetary rollers 3, which due to the parallelism of the axes la and 4a inevitably has the consequence that the falling lines FL1 and FL2 of the cone pulley tread 11a and the planetary roller tread, respectively 3a are oriented exactly parallel to each other. The falling lines FL1 and FL2 result in an imaginary section of the conical disk 11 or each planetary roller 3 through a sectional plane which passes through the longitudinal axis la of the transmission and the axis of rotation 4a of the planetary roller 3 in question. One of these sectional planes coincides with the plane of the drawing in FIG. 2. In the space between the conical disk 11 and each planetary roller 3 there is a double-conical intermediate roller 5, the first tapered running surface 5a of which rolls on the tapered running surface 3a of the associated planetary roller 3 and the second tapered running surface 5b of which rolls on the tapered running surface 11a of the conical disk 11. Each intermediate roller 5 is rotatable about an axis of rotation 6a and axially displaceably mounted on a bearing shaft 6, the left end of which is rigidly fixed in the planet carrier 2b. The axis of rotation 6a of each intermediate roller 5, which is identical to the longitudinal axis of the bearing shaft 6, is arranged at an angle with respect to the parallel falling lines FL1 and FL2 of the conical disk tread 11a and the planetary roller tread 3a, preferably in such a way that the axis of rotation 6a is in the drawn position of the intermediate roller 5 with the transmission longitudinal axis la and the falling line FL1 at point S1 and on the other hand intersects with the axis of rotation 4a of the associated planetary roller 3 and the falling line FL2 of the tread 3a of this assigned planetary roller 3 at point S2. This position shown in FIG. 2 represents the ideal position of the intermediate rollers 5, in which no drilling friction occurs on the linear friction contacts of the running surfaces 5a and 3a or 5b and 11a.

The length of the treads 5a, 5b and the diameter of each intermediate roller 5 are dimensioned such that the pressure field DF shown with double hatching in FIG. 2 is formed symmetrically to the center S3 of each intermediate roller 5. Thus, the pressure gradients p of the pressure field DF cannot exert a tilting moment on the intermediate roller 5 in question around its center point S3.

For the function of the roller transmission 1 shown in Figures 1 and 2, it is essential that the planet carrier 2b is continuously adjustable in the direction of the longitudinal axis la of the transmission with the aid of an only schematically illustrated actuator 2, the axis bores in which in the event of an adjustment 8th

Planet carriers 2b slide back and forth on the carrying axles 4 inserted therein, as indicated by the double arrow 2c in FIG. 2. Furthermore, when the planet carrier 2b is adjusted, the bearing shafts 6 of the intermediate rollers 5 fastened in the planet carrier 2b move correspondingly parallel to the longitudinal axis of the gearbox la. The center S3 of each intermediate roller 5 follows the intersection of the respective axis of rotation 6a with the central parallel 9 between the corresponding falling lines FL1 and FL2. Thus, when the planet carrier 2b is adjusted, each intermediate roller 5 moves along its position shaft 6.

In the inner end position of each intermediate roller 5 (with the shortest distance from the free end of the bearing shaft 6), the diameter ratio between the effective diameter of the conical disk 11 and the effective diameter of the individual planet rollers 3 is the smallest, so that when the drive shaft 7 is driven with a given drive speed the speed of the output shaft 8 is the smallest. Conversely, when each intermediate roller 5 is in its outer end position, in which the distance between the intermediate roller 5 and the free axial end of its bearing shaft 6 is greatest, the ratio between the effective diameter of the conical pulley 11 and the relevant planetary roller 3 largest, so that for a given input speed of the drive shaft 7, the output speed of the output shaft 8 is greatest. It can be seen that an infinitely variable adjustment of the intermediate rollers 5 along their axes of rotation 6a and thus an infinitely variable change in the gear ratio of the transmission 1 can be achieved by means of the infinitely variable adjustment of the planet carrier 2b.

In the first embodiment of the continuously variable roller transmission 1 described with reference to FIGS. 1 and 2, the directions of rotation of the drive shaft 7 and the output shaft 8 are opposite. In order to achieve the same directions of rotation of the input and output, the first embodiment required 9

Figures 1 and 2 only the output shaft 8 to be held, whereby the planet carrier 2a would rotate about the longitudinal axis la of the transmission in the same direction to the drive shaft 7. In the event of a rotational movement of the planet carrier 2a, the planet carrier 2b naturally also rotates in the same direction as the planet carrier 2a.

In the second embodiment of the roller transmission according to the invention illustrated with reference to FIGS. 3 to 7, only two transmission stages according to FIGS. 1 and 2 are coupled in mirror image to one another, the spur gears 3b and 8a being omitted. Instead, the planetary rollers 3 are double-tapered, which inevitably results from the mirror-symmetrical coupling of the two gear stages according to FIGS. 1 and 2. Furthermore, the support axles 4 of the planet rollers 3 are now mounted in both planet carriers 2a, 2b in a rotationally fixed and axially immovable manner. When the planet carrier 2a is displaced with the aid of the actuator 2 in the direction of the longitudinal axis la of the transmission, the other planet carrier 2b is therefore adjusted by the same amount and in the same direction, with the result that both gear stages of the embodiment according to FIGS. 3 and 4 are adjusted in opposite directions . This means, as from Fign. 5 and 6 it can be seen that in the left, drive-side gear stage, the intermediate rollers 5 are in their outer end position when the intermediate rollers 5 are in their inner end position in the right, output-side gear stage. Since the left gear stage is driven by the drive shaft 7 and the right gear stage is driven by the planet gears 3, the aforementioned opposite adjustment of both gear stages leads to a multiplication of the individual gear ratios of the gear stages in the same direction. The direction of rotation of the drive shaft 7 therefore coincides with the direction of rotation of the output shaft 8 in FIG. 4. 10

In a variant of the second embodiment according to FIG. 7, the axes of rotation 4a of the planetary rollers 3 can be pivoted from their position parallel to the longitudinal axis of the gearbox la by a maximum angle of + α and -α (corresponding to the axis positions 4al and 4a2 in FIG. 7). This takes place, for example, in that the planet carrier 2a is rotated by a corresponding angle of rotation (relative to the longitudinal axis la of the transmission) with respect to the other planet carrier 2b. The mounting of the support axles 4 must accordingly be designed so that they can move freely. As a result of the pivoting of the axes of rotation 4a, there is a corresponding oblique running between the running surfaces 3a of the planetary rollers 3 and the running surfaces 5a of the associated intermediate rollers 5. As a result of this oblique running, the intermediate rollers 5 move automatically and without effort along their axes of rotation 6 with a corresponding change in the gear ratio. As soon as the desired gear ratio is reached, the axes of rotation 4a of the planetary rollers 3 are moved back into the neutral position parallel to the longitudinal axis la of the gearbox. In this way, the displacement of the planet carrier 2a parallel to the longitudinal axis la of the transmission is facilitated, whereby considerably lower adjustment forces are required. According to FIG. 7, an actuating pinion 2 ′ meshing with the planet carrier 2a and having a coupled servomotor 2 ″ can be used as the actuator 2 for rotating the planet carrier 2a.

FIGS. 8 and 10 illustrate a third embodiment of the continuously variable roller transmission according to the invention. Compared to the first two embodiments, in the third embodiment, the planetary rollers 30 are not located on the outer circumference of the roller transmission 1 with axes of rotation parallel to the longitudinal axis of the transmission la, but are arranged in a star shape in the interior of the roller transmission 1 so that their axes of rotation 40a are oriented perpendicular to the longitudinal axis of the transmission la. The planetary rollers 30 are simple 11

Truncated cones are formed, the smaller diameter of each planetary roller 30 being located radially further outward with respect to the longitudinal axis of the gearbox la and the larger diameter of each planetary roller 30 being located radially further inside with respect to the longitudinal axis of the gearbox la. As a result of the star-shaped arrangement of the planetary rollers 30, the conical disks 100, 110 in the embodiment according to FIGS. 8 and 10 are designed in the form of hollow cones, so that their conical running surfaces 100a, 110a have the same inclination as the conical running surfaces 30a of the planetary rollers 30. This means that in the same way as in the first two embodiments, the falling lines FL1 and FL2 of the running surfaces 110a and 30a run exactly parallel to one another.

Due to the displacement of the planetary rollers 30 into the interior of the roller transmission 1 and the transformation of the conical disks 100, 110 into hollow cones, the bearing shafts 6 of the intermediate rollers 5 are fastened in the area of the transmission center. For this purpose, a carrier hub 20a or 20c is provided for each gear stage or group of intermediate rollers 5, which is arranged rotationally symmetrically about the longitudinal axis la of the gearbox. The two hubs 20a, 20c are rigidly connected to one another, for example with the aid of connecting bolts 21, as illustrated in FIGS. 9 and 13. The carrier hubs 20a, 20c, which are rigidly coupled to one another, can be displaced with the aid of the actuator 2 in the direction of the longitudinal axis la of the transmission, as indicated in FIG. 9 by the double arrows 2c. The carrier hub 20a is penetrated by the output shaft 80, which is rotationally coupled to the planet carrier 20b, which is located between the two carrier hubs 20a and 20c. In the case of FIGS. 9 and 13, the planet carrier 20b is penetrated by the connecting bolts 21 between the two carrier hubs 20a, 20b, so that the two carrier hubs 20a, 20c can move relative to the planet carrier 20b in the direction of the longitudinal axis of the gearbox la. The planet carrier 20b, for example pentagonal (FIG. 13), has bores in the same way as the planet carrier 2a in FIG 12

Support axes 40 for the planetary rollers 30 arranged in a star shape around the planet carrier 20b. The planet carrier 20b, which is coupled in a rotationally fixed manner to the output shaft 80, takes the connecting bolts 21 between the carrier hubs 20a, 20c with it when it rotates about the longitudinal axis la of the transmission, so that the carrier hubs 20a, 20c also rotate about the longitudinal axis of the transmission la in the direction of the output shaft 80. The drive shaft 70 is connected in a rotationally fixed manner to the rotating conical disk 110, while the other, fixed conical disk 100 is connected in a rotationally fixed manner to the transmission housing 12.

The conical disks 100, 110 have a central recess so that there is sufficient free space in the transmission 1 for the axial adjustment of the carrier hubs 20a, 20c.

With regard to the arrangement of the axes of rotation 6a of the intermediate rollers 5 with respect to the falling lines FLl and FL2 of the treads 110a and 30a, the same relationships and intersections S1 and S2 result in the embodiment according to FIGS. 8 and 10 as in the first and second embodiment of FIG Rolling gear according to the invention according to Figures 1 and 3. The same applies to the dimensioning of the length of the running surfaces 5a, 5b and the radius' of each intermediate roller 5 of the rolling gear according to Figures 8 and 10. The direction of rotation of the output shaft 80 is in the third embodiment according to Figures 8 and 10 the same as the direction of rotation of the drive shaft 70, ie there is a synchronism of drive and output. A reversal of the direction of rotation of the input and output can be achieved in the third embodiment according to FIGS. 8 and 10 in that the conical disk 100 is designed to be rotatable. The output torque is then taken from the rotatable conical disk 100, the output shaft 80 being held at the same time. 13

In the embodiment according to FIG. 10, the conical disk 110 is preloaded in the direction of the conical disk 100 by a disk spring 71 arranged rotationally symmetrically to the longitudinal axis la of the gearbox in order to ensure the greatest possible frictional force between all the rolling elements of the gearbox 1. The plate spring 71 is supported in one turn of the gear housing 12 against an axial bearing 72 of the conical disk 110. Instead of the plate spring 71, a pretensioning device which is dependent on the drive torque can also be provided, as is known per se from the prior art.

The two end positions of the third embodiment of the roller transmission 1 according to the invention are illustrated in FIGS. 11 and 12, FIG. 11 illustrating the case of the lowest output speed for a given input speed and FIG. 12 illustrating the case of the largest output speed for a given input speed.

A further possibility for adapting the frictional forces of the rolling gear to the output torque is indicated in FIG. 13. The output shaft 80 is in the form of a polygon in the region of the planet carrier 20b, the number of edge surfaces corresponding to the number of planet rollers 30. The support axis 40 of each planetary roller 30 is axially displaceably mounted within its associated bore in the planet carrier 20b and is supported with its spherically shaped free end on the associated edge surface of the output shaft 80, which is designed as a polygon. The supporting force between the free end of each support axis 40 and the associated edge surface of the output shaft 80 results from the contact pressure of the planet rollers 30. Furthermore, the torque of the output shaft 80 increases the wedge effect between the edge surfaces of the output 14

shaft 80 and the spherical free ends of the support axles 40, which in turn has the result that the existing supporting force and thus the contact pressure of the planetary rollers 30 against the intermediate rollers 5 is increased. A basic level of the contact pressure between the planet rollers 30 and the intermediate rollers 5 is provided by spring elements 31, which are each arranged between a collar 41 of the support shaft 40 and the relevant planet roller 30 and generate a corresponding preload on the planet rollers 30. At the same time, the spring elements 31 can compensate for irregularities in the contact pressures.

As can also be seen in FIG. 13, the axes of rotation 40a of the planetary rollers 30 can be rotated in an angular range of + α and -α relative to the neutral position in an analogous manner as in FIG. 7 in order to cause the planetary rollers 30 to skew on the intermediate rollers 5. In this way, in accordance with FIG. 7, the adjustment of the gear ratio of the transmission 1, the axial displacement of the carrier hubs 20a and 20c can be facilitated. The structural details for the adjustment of the axes of rotation of the planet rollers 30 are not shown in FIG.

In a similar way to the first two embodiments, it is also possible in the case of the third embodiment according to FIGS. 8 to 12 to use only half of the third embodiment constructed in mirror symmetry. This one-stage design is shown as the fourth embodiment of the invention in FIGS. 14 and 15. Analogously to FIG. 1, in this fourth embodiment of the roller transmission according to the invention, the conical planet rollers 30 are coupled to the output shaft 80 via rigidly coupled bevel gear teeth 30b, 80a. 15

A fifth embodiment of the roller transmission according to the invention is illustrated with reference to FIGS. 16 and 17. In contrast to the third embodiment according to FIGS. 8 and 10, in the fifth embodiment according to FIGS. 16 and 17 the running surfaces 100a and 110a of the conical disks 100 and 110 are not designed as straight conical surfaces but as curved conical surfaces. In the same way, the running surfaces 30a of the planetary rollers 30 are designed as curved conical surfaces. The running surfaces 5a and 5b of each double-cone-shaped intermediate roller 5 are also designed as curved conical surfaces. As a result of the curvature of all conical surfaces, the center point S3, which runs concentrically to the curved surfaces of the assigned planetary roller 30 and the associated conical disk 100 or 110, moves on a circular path 90. So that each intermediate roller 5 can follow the circular path 90, its bearing shaft 6 is in the associated carrier hub 20a or 20c pivotally mounted.

In order to achieve minimal drilling friction, it is necessary in the embodiment according to FIGS. 16 and 17 that the intersections S1 and S2 of the axis of rotation 6a of each intermediate roller 5 with the falling lines FL1 and FL2 on the one hand on the longitudinal axis of the gearbox la and on the other hand on the axis of rotation 40a of the associated planetary roller 30 come to lie approximately within their swivel range. This results in a design limitation of the number of planetary rollers 30 to two diametrically opposed planetary rollers 30. If the number of planetary rollers 30 according to FIGS. 16 and 17 can be increased per se, a higher drilling friction must be accepted. Compared to the embodiments with straight conical surfaces, the low drilling friction is present in the embodiment according to FIGS. 16 and 17 in a substantially constant manner over the entire adjustment range. Furthermore, 16

initiate the adjustment forces via a pivoting movement of the bearing shafts 6 of the intermediate rollers 5, which requires lower adjustment forces.

A sixth embodiment of the roller transmission according to the invention is illustrated with reference to FIGS. 18 and 19. In contrast to the second embodiment according to FIGS. 3 and 4, in the sixth embodiment according to FIGS. 18 and 19 the treads 10a, 11a of the conical disks 10, 11 and the treads 3a of the external planet rollers 3 are not convex, but concave. The position and attachment of the bearing shafts 6 for the intermediate rollers 5 is adapted in accordance with this change in the tread inclination in such a way that the bearing shafts are no longer fastened in the planet carriers 2a, 2b as in FIG. 3, but now in the supporting axes 4 of the planetary rollers 3 . The support axes 4 in FIGS. 18 and 19 have a substantially larger diameter than the support axes 4 in FIGS. 3 and 4. The planetary rollers 3 in FIGS. 18 and 19 likewise have a substantially larger diameter than the planetary rollers 3 in FIGS. 3 and 4 and are supported with respect to the supporting axis 4 with their outer lateral surface on a central rolling element 22 which is arranged coaxially to the longitudinal axis of the gearbox la. For better guidance of the planetary rollers 3 on the central rolling element 22, the outer lateral surfaces of the planetary rollers 3 have a graduated recess in which the central rolling element 22 engages with its outer profile designed in the opposite direction to the profile of the recess. The conical disks 10, 11 are connected to one another via a shaft 7, which is arranged coaxially with the longitudinal axis la of the transmission and carries the rolling element 22. The speed difference between the conical disks 10 and 11 is compensated for by a compensating bearing 23, which in the example shown is arranged between the right axial end of the shaft 7 and the conical disk 10. 17

The advantage of the embodiment according to FIGS. 18 and 19 consists in the relatively large effective diameters of the conical disks 10, 11 and the planetary rollers 3. In addition, the concave treads 10a, 11a and 3a result in a broad line of contact with the intermediate rollers 5, which results in an increased Performance of the transmission results.

Claims

18P ATENT REQUESTS

1. Infinitely adjustable roller gear, characterized by the following features:

(a) at least one conical disk (10, 11; 100, 110) with a conical running surface (10a, 11a; 100a, 110a), which is arranged rotationally symmetrically with respect to the longitudinal axis of the gearbox (la);

(b) a group of rotatably mounted planet rollers (3; 30) which are arranged rotationally symmetrically with respect to the longitudinal axis of the gear (la) and each have at least one tapered running surface (3a; 30a);

(c) at least one group of double-cone-shaped intermediate rollers (5), which are arranged in a rotationally symmetrical manner between the planet rollers (3; 30) and the conical disk (10, 11; 100, 110) with respect to the longitudinal axis (la) of the gear, c2) each rotatable and are mounted for longitudinal displacement, and c3) each have two separate tapered running surfaces (5a, 5b); 19

wherein the mutually facing treads (10a, 11a; 100a, 110a; 3a; 30a) of the conical disc (10, 11; 100, 110) and the planetary rollers (3; 30) are designed in such a way that the cutting plane through the longitudinal axis of the gearbox (la ) running falling lines (FLl, FL2) of these treads are oriented parallel to one another,

wherein each intermediate roller (5) rolls with its first tapered tread (5a) on an associated planetary roller (3; 30) and with its second tapered tread (5b) on the tapered tread (10a, 11a; 100a, 110a) of the conical disk (10 , 11; 100, 110) rolls,

wherein the axes of rotation (6a) of the intermediate rollers (5) with respect to the falling lines (FLl, FL2) of respectively assigned tapered running surfaces (10a, 11a; 110a, 110a; 3a; 30a) of the conical disk (10, 11; 100, 110) and the Planetary rollers (3; 30) are arranged at an angle, and

wherein the axes of rotation (6a) of all intermediate rollers (5) of each intermediate roller group together with respect to the conical disk (10, 11; 100, 110) and the planetary rollers (3; 30) can be displaced essentially parallel to the longitudinal axis of the gearbox (la), such that the intermediate rollers (5) each intermediate roller group with a displacement of their axes of rotation (6a) in the direction of the longitudinal axis of the gearbox (la) forcibly moved along their axes of rotation (6a) while changing the gear ratio of the roller transmission. 20th

2. Infinitely adjustable roller gear according to claim 1, characterized in that in each of a planetary roller (3; 30) and an intermediate roller (5) existing roller pairing a position is adjustable in which the axis of rotation (4a; 40a) of the planetary roller (3rd ; 30) and the axis of rotation (6a) of the intermediate roller (5) with the fall line (FL2) of the associated running surface (3a; 30a) of the planetary roller (3; 30) intersects at a common point (S2).

3. Infinitely variable rolling gear according to claim 1 or 2, characterized in that in each of a planetary roller (3; 30) and an intermediate roller (5) existing roller pairing a position is adjustable in which the coinciding with the longitudinal axis of the gearbox (la), common longitudinal axis of the conical disk (10, 11; 100, 110), the axis of rotation (6a) of the intermediate roller (5) and the falling line (FLl) of the associated tread (10a, 11a; 100a, 110a) of a conical disk (10, 11; 100 , 110) intersects at a common point (S1).

4. Infinitely adjustable rolling gear according to one of claims 1 to 3, characterized in that all tapered treads are straight tapered surfaces, with a linear contact between the rolling surfaces rolling. 21

5. Infinitely adjustable rolling gear according to one of claims 1 to 3, characterized in that all tapered running surfaces are curved conical surfaces, with a linear contact between the rolling surfaces rolling on one another.

6. Infinitely adjustable roller gear according to one of claims 1 to 5, characterized in that the width and the taper angle of the first and second running surfaces (5a, 5b) of each intermediate roller (5) and the diameter thereof are selected such that the first and second Treads (5 a, 5b) are essentially mirror-symmetrical with respect to a normal (normal) on the adjacent running surfaces of the associated planetary roller (3; 30) and the conical disk (10, 11; 100, 110), which through the center point (S3) of the Intermediate roller (5) runs.

7. Infinitely adjustable rolling gear according to one of claims 1 to 6, characterized in that each planetary roller (3) with its axis of rotation (4a) parallel to the longitudinal axis of the gearbox (la) and the coincident longitudinal axis of the conical disk (10, 11; 100, 110) is arranged.

8. Infinitely adjustable roller gear according to one of claims 1 to 7, characterized in that only a single drive-side conical disk (11; Figures 1 and 2) is provided that the planetary rollers (3) are each coupled to a spur gear (3b) , 22

and that all spur gears (3b) of the planet rollers (3) mesh with a ring gear (8a).

9. Infinitely variable roller gear according to one of claims 1 to 7, characterized in that two conical disks (10, 11; 100, 110) are provided at a distance from each other, that each planetary roller (3) is double-conical with two tapered running surfaces (3a) and that two groups of intermediate rollers (5) are provided, the intermediate rollers (5) of one group on the first tapered running surfaces (3a) of the planetary rollers (3) and the intermediate rollers (5) of the other group on the second tapered running surfaces (3a) unroll the planet rollers (3).

10. Infinitely adjustable roller gear according to one of claims 1 to 9, characterized in that the planet rollers (3) in two annular, opposite planet carriers (2a, 2b) are mounted.

11. Infinitely adjustable rolling gear according to claim 10, characterized in that the bearing shafts (6) of all intermediate rollers (5) of an intermediate roller group in the adjacent planet carrier (2a or 2b) are fixedly mounted.

12. Infinitely adjustable rolling gear according to claim 10 or 11, characterized in that a planet carrier (for example 2a) relative to the other planet carrier (for example 2b) can be rotated by an adjusting angle with respect to the longitudinal axis of the transmission (la) 23

is that when the one planet carrier (2a) is expanded, the axes of rotation (4a) of the planetary rollers (3) can be pivoted by a slip angle (+ α, -α) relative to a neutral position.

13. Infinitely adjustable roller gear according to one of claims 1 to 6, characterized in that each planetary roller (30) is designed in the form of a simple truncated cone and with its axis of rotation (40a) perpendicular to the longitudinal axis of the gearbox (la) and the longitudinal axis of the conical disc (10, 11; 100, 110) is arranged.

14. Infinitely adjustable roller gear according to one of claims 1 to 6 and 13, characterized in that only a single drive-side conical disk (110; Figures 14 and 15) is provided, that the planetary rollers (30) are each coupled to a bevel gear (30b) , and that all bevel gears (30b) of the planetary rollers (30) mesh with a central bevel gear (80a).

15. Infinitely adjustable roller gear according to one of claims 1 to 6 and 13, characterized in that two conical disks (10, 11; 100, 110) are provided at a distance, and that two groups of intermediate rollers (5) are provided, wherein Roll the intermediate rollers (5) of both groups on the tapered running surfaces (30a) of the planetary rollers (30). 24

16. Infinitely adjustable roller gear according to claim 15, characterized in that the bearing shafts (6) of all, on a common conical disk (100 or 110) rolling intermediate rollers (5) distributed in a star shape are attached to a common carrier hub (20a or 20c) , wherein the two carrier hubs (20a and 20c) are arranged rotationally symmetrically with respect to the longitudinal axis of the gearbox (la) and are rigidly connected to one another, and that all axle journals (40) of the planetary rollers (30) are distributed in a star shape in a common, central carrier slide (20b) are mounted, which is slidably mounted on the fixed connection (21) between the two axially displaceable carrier hubs (20a or 20c).

17. Infinitely adjustable roller gear according to claim 16, characterized in that the carrier hub (20b) is coupled to the output shaft (80) and that the output-side conical disk (100) is fixed.

18. Infinitely variable rolling gear according to claim 16, characterized in that the carrier hub (20b) is fixed and that the output-side conical disc (100) is coupled to the output shaft (80).

19. Infinitely adjustable rolling gear according to claim 16, characterized in that the carrier slide (20b) relative to the carrier hubs (20a, 20c) by an adjusting angle (+ α, -α) with respect to the transmission longitudinal axis (la) is rotatable. 25th

20. Infinitely variable rolling gear according to claim 16, characterized in that the supporting axes (40) of the planetary rollers (30) in the carrier slide (20b) are axially displaceably supported and are supported with their axial ends on edge surfaces of the polygonal output shaft (80).

21. Infinitely adjustable rolling gear according to claims 5 and 16, characterized in that when all tapered running surfaces (5a, 5b, 30a, 100a, 110a) are formed as curved tapered surfaces, the bearing shafts (6) of the intermediate rollers (5) are articulated on the carrier hubs ( 20a, 20c) are stored.

22. Infinitely adjustable roller gear according to one of claims 5 1 to 21, characterized in that the running surfaces (10a, 11a or 3a) of the conical disks (10, 11) and planetary rollers (3) are concave and that the bearing shafts (6) the intermediate rollers (5) are fastened to the bearing shafts (4) of the planet rollers (3) (FIGS. 18 and 19).

23. Infinitely adjustable rolling gear according to claim 22, characterized in that each planetary roller (3) is mounted with its outer lateral surface on a central rolling element (22) in a rolling manner.

24. Infinitely adjustable rolling gear according to claim 23, characterized in that the outer lateral surface of each planetary roller (3) has a recess in which the central rolling element (22) engages. 26

25. Infinitely adjustable roller gear according to one of the claims

22 to 24, characterized in that. the conical disks (10, 11) are connected to one another via a central shaft (7) and a compensating bearing (23).