WO1999039115A1 - Transmission a roulement a reglage continu - Google Patents

Transmission a roulement a reglage continu Download PDF

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Publication number
WO1999039115A1
WO1999039115A1 PCT/EP1999/000625 EP9900625W WO9939115A1 WO 1999039115 A1 WO1999039115 A1 WO 1999039115A1 EP 9900625 W EP9900625 W EP 9900625W WO 9939115 A1 WO9939115 A1 WO 9939115A1
Authority
WO
WIPO (PCT)
Prior art keywords
rollers
roller
planetary
gear according
longitudinal axis
Prior art date
Application number
PCT/EP1999/000625
Other languages
German (de)
English (en)
Inventor
Georg Albersinger
Original Assignee
Georg Albersinger
Priority date (The priority date is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the date listed.)
Filing date
Publication date
Application filed by Georg Albersinger filed Critical Georg Albersinger
Publication of WO1999039115A1 publication Critical patent/WO1999039115A1/fr

Links

Classifications

    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16H—GEARING
    • F16H15/00—Gearings for conveying rotary motion with variable gear ratio, or for reversing rotary motion, by friction between rotary members
    • F16H15/02—Gearings for conveying rotary motion with variable gear ratio, or for reversing rotary motion, by friction between rotary members without members having orbital motion
    • F16H15/04—Gearings providing a continuous range of gear ratios
    • F16H15/06—Gearings providing a continuous range of gear ratios in which a member A of uniform effective diameter mounted on a shaft may co-operate with different parts of a member B
    • F16H15/16—Gearings providing a continuous range of gear ratios in which a member A of uniform effective diameter mounted on a shaft may co-operate with different parts of a member B in which the member B has a conical friction surface
    • F16H15/18—Gearings providing a continuous range of gear ratios in which a member A of uniform effective diameter mounted on a shaft may co-operate with different parts of a member B in which the member B has a conical friction surface externally
    • F16H15/20—Gearings providing a continuous range of gear ratios in which a member A of uniform effective diameter mounted on a shaft may co-operate with different parts of a member B in which the member B has a conical friction surface externally co-operating with the outer rim of the member A, which is perpendicular or nearly perpendicular to the friction surface of the member B
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16H—GEARING
    • F16H15/00—Gearings for conveying rotary motion with variable gear ratio, or for reversing rotary motion, by friction between rotary members
    • F16H15/48—Gearings for conveying rotary motion with variable gear ratio, or for reversing rotary motion, by friction between rotary members with members having orbital motion
    • F16H15/50—Gearings providing a continuous range of gear ratios

Definitions

  • 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.
  • 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.
  • 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.
  • a roller gear should preferably have a high power density so that it can be used in vehicle construction.
  • the invention is based on the consideration that low-friction transmission with adjustable transmission ratio can be achieved by tapered rolling elements.
  • 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.
  • FIG. 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
  • 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,
  • FIG. 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.
  • 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.
  • 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.
  • a central drive shaft 7 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.
  • 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.
  • each intermediate roller 5 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.
  • 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.
  • 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.
  • each intermediate roller 5 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.
  • 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.
  • the planet carrier 2b naturally also rotates in the same direction as the planet carrier 2a.
  • 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 + ⁇ and - ⁇ (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.
  • FIGS. 8 and 10 illustrate a third embodiment of the continuously variable roller transmission according to the invention.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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
  • 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.
  • 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
  • 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.
  • a pretensioning device which is dependent on the drive torque can also be provided, as is known per se from the prior art.
  • FIGS. 11 and 12 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.
  • 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.
  • the torque of the output shaft 80 increases the wedge effect between the edge surfaces of the output 14
  • 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.
  • the axes of rotation 40a of the planetary rollers 30 can be rotated in an angular range of + ⁇ and - ⁇ 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.
  • 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.
  • FIGS. 16 and 17 A fifth embodiment of the roller transmission according to the invention is illustrated with reference 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.
  • 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.
  • 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.
  • FIGS. 18 and 19 A sixth embodiment of the roller transmission according to the invention is illustrated with reference 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 planetary rollers 3 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.
  • 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.
  • 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.

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  • Engineering & Computer Science (AREA)
  • General Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • Friction Gearing (AREA)

Abstract

L'invention concerne une transmission à roulement à réglage continu, s'utilisant de préférence dans la construction automobile, comportant au moins une poulie conique (10, 11; 100, 110), un groupe de rouleaux planétaires (3; 30) montés rotatifs et présentant au moins une surface de roulement conique (3a; 30a), et au moins un groupe de rouleaux intermédiaires (5) biconiques montés rotatifs, qui peuvent se déplacer longitudinalement et sont positionnésentre les rouleaux planétaires (3; 30) et la poulie conique (10,11; 100, 110). Les poulies coniques, les rouleaux planétaires etles rouleaux intermédiaires sont disposés à symétrie de révolution par rapport à l'axe longitudinal (1a) de la transmission. Les surfaces de roulement (10a, 11a; 110a; 3a; 30a) de la poulie conique (10, 11; 100, 110) et des rouleaux planétaires (3; 30), situées en regard les unes des autres, sont configurées de telle manière que les lignes de plan de coupe (FL1, FL2) desdites surfaces deroulement, passant par l'axe longitudinal (1a) de la transmission dans des plans de coupe, sont parallèles les unes aux autres. Les axes de rotation (6a) des rouleaux intermédiaires (5) sont positionnés en biais par rapport aux lignes de plan de coupe (FL1, FL2) des surfaces de roulement coniques correspondantes (10a, 11a;
PCT/EP1999/000625 1998-02-02 1999-02-01 Transmission a roulement a reglage continu WO1999039115A1 (fr)

Applications Claiming Priority (2)

Application Number Priority Date Filing Date Title
DE19804011.3 1998-02-02
DE1998104011 DE19804011A1 (de) 1998-02-02 1998-02-02 Stufenlos verstellbares Wälzgetriebe

Publications (1)

Publication Number Publication Date
WO1999039115A1 true WO1999039115A1 (fr) 1999-08-05

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Family Applications (1)

Application Number Title Priority Date Filing Date
PCT/EP1999/000625 WO1999039115A1 (fr) 1998-02-02 1999-02-01 Transmission a roulement a reglage continu

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DE (1) DE19804011A1 (fr)
WO (1) WO1999039115A1 (fr)

Cited By (2)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
ES2353539A1 (es) * 2008-10-29 2011-03-03 Universidad De Leon Caja de cambios de variación continua.
CN102927227A (zh) * 2012-09-01 2013-02-13 郭克亚 一种长锥滚锥式无级变速传动机构

Families Citing this family (4)

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Publication number Priority date Publication date Assignee Title
CZ305972B6 (cs) * 2003-07-23 2016-06-01 Jasoň Hampl Planetový spojitě měnitelný převod
DE102004002284A1 (de) * 2004-01-16 2005-08-04 Zf Friedrichshafen Ag Stufenlosgetriebe
DE502005005158D1 (de) 2005-01-15 2008-10-02 Luk Lamellen & Kupplungsbau Variator für ein stufenloses getriebe
CZ2010137A3 (cs) * 2010-02-24 2011-11-23 Ceská zemedelská univerzita v Praze Trecí sférický planetový variátor otácek

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Publication number Priority date Publication date Assignee Title
DE1007589B (de) * 1953-03-25 1957-05-02 Daimler Benz Ag Waelzkoerper-Reibgetriebe
GB2006354A (en) * 1977-10-14 1979-05-02 Patent Concern Nv Torque converter
US4183253A (en) * 1977-02-07 1980-01-15 Domenic Borello Dual surface angular power transfer wheel traction nipping device
EP0010329A1 (fr) * 1978-10-13 1980-04-30 Patent Concern N.V. Convertisseur de couple

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Publication number Priority date Publication date Assignee Title
US1844239A (en) * 1929-03-14 1932-02-09 Boehme Inc H O Power transmission
AT179455B (de) * 1953-05-13 1954-09-10 Margarete Prieschink Stufenlos regelbares Planetenreibradwechsel- und -wendegetriebe
FR1376401A (fr) * 1963-12-05 1964-10-23 Fabrications Unicum Soc D Perfectionnements au dispositif de réglage des variateurs de vitesse à friction enparticulier
DE2219879A1 (de) * 1972-04-22 1973-10-25 Heinrich Ruebel Kugelreibgetriebe mit umlaufenden kegelrollen
US5545101A (en) * 1993-07-20 1996-08-13 Ntn Corporation Friction type continuously variable transmission
NL9301646A (nl) * 1993-09-23 1995-04-18 Doornes Transmissie Bv Traploos instelbare transmissie.

Patent Citations (4)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
DE1007589B (de) * 1953-03-25 1957-05-02 Daimler Benz Ag Waelzkoerper-Reibgetriebe
US4183253A (en) * 1977-02-07 1980-01-15 Domenic Borello Dual surface angular power transfer wheel traction nipping device
GB2006354A (en) * 1977-10-14 1979-05-02 Patent Concern Nv Torque converter
EP0010329A1 (fr) * 1978-10-13 1980-04-30 Patent Concern N.V. Convertisseur de couple

Cited By (2)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
ES2353539A1 (es) * 2008-10-29 2011-03-03 Universidad De Leon Caja de cambios de variación continua.
CN102927227A (zh) * 2012-09-01 2013-02-13 郭克亚 一种长锥滚锥式无级变速传动机构

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