CN100470091C - continuously variable transmission - Google Patents

Abstract

A variable speed transmission having a plurality of tilting balls (1) and opposing input (34) and output discs (101) is shown and described, the transmission having an infinite number of speed combinations in its ratio range. The use of a planetary gear set allows the minimum speed to be reversed (in reverse), and the unique geometry of the transmission allows all of the power paths to be coaxial, thereby reducing the overall size and complexity of the transmission as compared to transmissions that achieve similar ratio ranges.

Description

Continuously variable transmission

technical field

The field of the invention relates generally to transmissions, and more particularly the invention relates to continuously variable transmissions.

Background technique

In order to provide continuously variable transmissions, various traction roller drives have been developed that transmit power through traction rollers supported in a frame between torque input and output discs. In such transmission equipment, the traction roller is mounted on a support structure which, when pivoted, engages the traction roller with a torque disc in a circle whose radius varies according to the desired transmission rate.

However, these traditional solutions have limitations. For example, in one solution, a drive axle for a vehicle with a continuously variable transmission is disclosed. This method teaches the use of two iris plates (one at each end of the pulling rolls) to tilt the axis of rotation of each pulling roll. However, using an iris plate is very complicated because of the large number of components required to adjust the iris plate during the switching of transmission equipment. A further disadvantage of this transport equipment is that it has guide rings which are arranged so that they are very stationary relative to each pulling roller. Since the guide rings are fixed, it is difficult to move the axis of rotation of each pulling roller.

A modification to this earlier design included a shaft about which the input and output disks rotated. An input disc and an output disc are mounted in the shaft and contact a plurality of balls disposed equidistantly and axially around the shaft. These balls are in frictional contact with the input and output discs and transfer power from the input discs to the output discs. An idler located over the center of the shaft and between the balls exerts force to keep the balls apart, thereby bringing the input and output discs into frictional contact. A major limitation of this design is the lack of means to generate and adequately control the axial force as conventional joint pressure to keep the input and output discs in sufficient contact with the balls as the speed of the transmission equipment changes . Since the roller traction continuously variable transmission requires more axial force at the lowest speed to prevent the driving and driven rotating elements from slipping on the friction balls with changing speed, at high speed and 1:1 ratio when the input and output speeds are equal Ratio to apply excessive force. This excessive axial force reduces efficiency and causes the transmission to fail sooner than when the proper amount of force is applied for any particular gear ratio. The excessive force also makes switching gears more difficult.

Accordingly, there is a need for a continuously variable transmission having an improved axial load generating system that varies the force generated as a function of the transmission ratio.

Contents of the invention

The systems and methods shown and described herein have several features, no single one of which is representative of its desirable attributes. The following description is by no means limiting the field presented, but its very important features will now be briefly described. After this discussion, and particularly after reading the heading "Detailed Description of the Preferred Embodiments," one should understand how the features of the systems and methods of the present invention provide advantages over conventional systems and methods.

In a first aspect, a variable speed transmission is disclosed comprising: a longitudinal axis; a plurality of balls radially distributed about the longitudinal axis, each of the balls having a tiltable axis and extending about the the shaft rotates; a rotatable input disc disposed adjacent to and in contact with each of the balls; a rotatable output disc disposed adjacent to the balls and in contact with the input disc opposite and in contact with each of said balls; a rotatable idler having a substantially constant outer diameter coaxial about said longitudinal axis and disposed radially inwardly of each of said balls and a planetary gear set coaxially mounted about the longitudinal axis of the transmission.

Also disclosed are embodiments of the variable speed transmission wherein the balls aggregate torque components transmitted from at least two power paths provided by the planetary gearset, wherein the at least two power paths are coaxial. In another embodiment, at least one of said idler gear and said output plate provides a torque input to said planetary gear set.

In another aspect, a variable speed transmission is disclosed wherein said planetary gearset further comprises: a ring gear coaxially mounted about said longitudinal axis and having radially inwardly disposed teeth; planetary gears coaxially distributed about said longitudinal axis within said ring gear and engaged with said ring gear, each of said planetary gears having a respective planetary axis and rotating about said planetary axis, wherein , the planetary axis is radially away from the longitudinal axis; a plurality of planetary gear shafts, each of the planetary gears has one of the planetary gear shafts, the planetary gears rotate around the planetary gear shaft; the sun gear, with axially mounted about said longitudinal axis and radially within and engaged with each of said plurality of planet gears; and a planet gear carrier coaxially mounted about said longitudinal axis and adapted to support The planet gear shaft and determine the position of the planet gear shaft.

Some of these embodiments further include a retainer adapted to align the tiltable axis of the ball, the retainer also adapted to maintain the angular and radial position of the ball. In some embodiments, the planet gear carrier is applied with input torque and is coupled to the input disc, the sun gear is coupled to the retainer, the ring gear is fixed and cannot rotate through the An output disc provides output torque from the transmission.

In another aspect, there is disclosed an axial force generating device usable with embodiments disclosed herein adapted to generate an increased force between the input disc, the balls, the idler, and the output disc. Axial force of traction. In some embodiments, the amount of axial force generated by the axial force generator is a function of the gear ratio of the transmission. In other embodiments, each of the input disc, the balls, the output disc and the idler has a contact surface coated with a friction increasing paint. Coatings of certain embodiments are ceramics or cermets. In other embodiments, the coating is selected from silicon nitride, silicon carbide, electroless nickel, electroplated nickel or any combination thereof.

In another aspect, a variable speed transmission is disclosed, comprising: a longitudinal axis; a plurality of balls radially distributed about the longitudinal axis, each of the balls having a tiltable axis and extending about the the shaft rotates; a rotatable input disc disposed adjacent to and in contact with each of the balls; a fixed output disc disposed adjacent to the balls and opposite the input disc , and in contact with each of the balls; a rotatable idler wheel having a substantially constant outer diameter and radially disposed in and in contact with each of the balls; a retainer, which is adapted to maintain the radial position and axial alignment of said balls and said retainer is rotatable about said longitudinal axis; and an idler shaft connected to said idler adapted to receive torque from said idler output and transmits the torque output from the transmission.

In another aspect, a variable speed transmission is disclosed comprising: first and second sets of balls radially distributed about said longitudinal axis; rotatable first and second input discs; an input shaft, coaxial with said longitudinal axis and connected to said first and second input discs; a rotatable output disc disposed between said first set of balls and said second set of balls and connected to said first set of balls contacting each of said first set of balls and said second set of balls; a generally cylindrical first idler wheel disposed radially inwardly of each of said first set of balls, and and a generally cylindrical second idler wheel disposed radially inwardly of each of said second set of balls and in contact therewith.

For use with many of the embodiments described herein, an axial force generator adapted to apply an axial force to increase the input disc, the output disc and contact force between said plurality of speed adjusters, said axial force generator further comprising: a bearing disc coaxial with said longitudinal axis and rotatable about said bearing disc having an outer diameter and an inner diameter , and having a threaded hole formed in its inner diameter; a plurality of outer peripheral ramps attached to the first side of the bearing disc proximate its outer diameter; a plurality of bearings adapted to engage the ramps of the plurality of bearing discs a plurality of input disc peripheral ramps mounted on said input disc on opposite sides of said speed regulator and adapted to engage said bearings; a generally cylindrical screw coaxial with said longitudinal axis and rotatable thereabout, the screw having male threads formed along its outer surface adapted to engage the threaded holes of the bearing disc; a plurality of central screw ramps attached to the screw; and a plurality of A central input disc ramp attached to the input disc and adapted to engage the plurality of central screw ramps.

In another aspect, a support holder for supporting and setting the position of a plurality of speed-adjustable tiltable balls in a rolling traction transmission using an input on either side of the plurality of balls is disclosed. and an output disc, the holder comprising first and second flat support discs, each of which is generally a disc having a plurality of slots extending radially inwardly from the outer edge, each a slot having two sides; a plurality of flat support spacers extending between said first and second support plates, each of said spacers having a front side, a rear side, a first end and a second end; wherein each of said first and second ends has a mounting surface, each mounting surface has a curved surface, said spacer is disposed at an angle around said support plate and between said slots in said support plate , so that the curved surface is aligned with the sides of the groove.

In another aspect, a support leg for a ratio changing mechanism is disclosed that changes the transmission ratio by inclining the axis of rotation of the balls forming the ratio in a rolling traction transmission, comprising: an elongated a main body; a shaft connection end; a cam end opposite the shaft connection end; a front side facing the ball and a rear side facing away from the ball; and a central support between the shaft connection end and the cam end part; wherein the shaft connection end has a hole formed adapted to receive the shaft therethrough, a convexly curved cam surface is formed on the front side of the cam end and is adapted to help control the hole arrangement.

Another aspect discloses a liquid-suction ball for use in a variable speed rolling traction drive using a plurality of balls rotatable about their respective tiltable axes, positioned at the an input disc on one side of and in contact with each of the plurality of spheres, and an output disc on the other side of and in contact with each of the plurality of spheres, the liquid suction spheres comprising: spherical spheres passing through The diameter of the sphere is formed with a hole forming through the cylindrical inner surface of the sphere; and at least one helical groove formed in the inner surface of the sphere and extending through the sphere.

In another aspect, a liquid suction shaft for use in a variable speed rolling traction drive using a plurality of spheres each having a shaft within the plurality of spheres is disclosed an input disk of said spheres on one side of each of said plurality of spheres and in contact therewith, and an output disk on the other side of and in contact with each of said plurality of spheres, the respective axes of which are formed by passing through said spheres Formed by a diameter hole formed by the sphere, the liquid pumping shaft includes a generally cylindrical shaft having a diameter smaller than the diameter of the hole passing through the sphere and having a first end, a second end, and an intermediate region, wherein , said first end and said second end extending from opposite sides of said sphere when said shaft is properly located within the bore of the respective sphere to which it belongs, said intermediate region being located within said sphere; and At least one helical groove is formed on the outer surface of the shaft, wherein the helical groove begins at a location external to the sphere and extends into at least a portion of the intermediate region.

In another aspect, a switching device is disclosed for a variable speed rolling traction drive having a longitudinal axis and utilizing a plurality of surrounding tilting balls, each opposite side of which is in contact with an input disc and an output disc, thereby controlling the transmission ratio of said transmission, said switching means comprising:

a tubular drive shaft disposed along said longitudinal axis; a plurality of ball shafts each extending through a hole formed through a respective one of said plurality of balls and forming a tiltable shaft of said respective ball, said ball rotates about said tiltable shaft, each ball shaft has two ends each extending from said ball; a plurality of legs each connected to said end of said ball shaft each of which is connected, the legs extending radially inwardly towards the drive shaft; an idler, having a substantially constant outer diameter, coaxially arranged around the drive shaft and radially located at the inward position of each of the balls and in contact with it; two disc-shaped switching guides, one at each end of the idler wheel, one of the switching guides each having a flat side facing said idler and a convex side facing away from said idler, wherein, on a respective side of said ball, said switching guides extend radially to contact all respective legs; a plurality of rolling pulleys, one for each leg, each attached to the side of its respective leg facing away from the ball; a generally cylindrical pulley table extending axially from the at least one of the switching guides extends; a plurality of guide pulleys, one for each rolling pulley, radially disposed about and attached to the pulley table; and a flexible tether , having first and second ends, the first end extending through the shaft and out of a slot formed in the shaft proximate the pulley block, the tether The first end is further wound on each of the rolling pulleys and each of the guide pulleys, wherein the second end extends from the shaft and reaches the switching device, and the guide pulleys are respectively mounted on one or more rings The second end pulls on each of the rolling pulleys to switch the transmission when the tether is pulled by the switching means to maintain alignment of each guide pulley with its respective rolling pulley.

In another aspect, a switching device for a variable speed rolling traction drive having a longitudinal axis and utilizing a plurality of tilted balls each having a ball radii from the centers of the respective balls, thereby controlling the gear ratio of the transmission, the switching means comprising: a plurality of ball shafts each extending through a bore formed through a corresponding one of the plurality of balls , and form the tiltable shafts of said respective balls, each ball shaft having two ends each extending from said ball; a plurality of legs each connected to said ball shaft Each of the ends is connected, the legs extending radially inwardly towards the drive shaft; a generally cylindrical idler wheel having a substantially constant radius, the idler wheel being coaxially disposed and having a diameter of each of the balls is provided at an inward position and in contact with it; first and second disk-shaped switching guides, each end of the idler wheel has one of the switching guides, the switching Each of the guides has a flat side facing the idler and a convex side facing away from the idler, wherein, on the respective side of the ball, the switching guides extend radially with respect to all of the respective the legs are all in contact; and a plurality of guide wheels each having a guide wheel radius, one guide wheel for each leg, each guide wheel being rotatably mounted at the radially inward end of its respective leg, wherein , the shape of the convex surface is determined by a set of two-dimensional coordinates whose origin is centered at the intersection of the longitudinal axis and a line drawn through the centers of any two diametrically opposed spheres, where it is assumed that at said convex surface is substantially tangent to said guide wheel at the point of contact between said guide wheel surface and said switching guide surface, said coordinates denote the position of said contact point as said idler wheel and said A function of the axial movement of the switching guide described above.

In another embodiment, an automobile is disclosed, comprising: an engine;

powertrain; and

Variable speed transmission, including:

a longitudinal axis; a plurality of balls radially distributed about said longitudinal axis, each ball having a tiltable axis and rotating about said ball; a rotatable input disc disposed adjacent to said ball and connected to contacting each of the balls; a rotatable output disc disposed adjacent to the balls and opposite the input disc and in contact with each of the balls; a rotatable idler having a substantially constant said longitudinal axis is coaxial with an outer diameter and is disposed radially inwardly of and in contact with each of said balls; and a planetary gear set coaxially mounted to said longitudinal direction of said transmission around the axis.

These and other modifications will become apparent to those skilled in the art after reading the following detailed description and the accompanying figures.

Description of drawings

Figure 1 is a cutaway side view of an embodiment of the transmission switched to high;

Figure 2 is a cutaway side view of an embodiment of the transmission switched to low;

Figure 3 is a partial cross-sectional view of the transmission taken from line III-III in Figure 1;

Figure 4 is a schematic cross-sectional side view of the idler and ramp fitting of the transmission of Figure 1;

Fig. 5 is a schematic perspective view of a ball fitting of the transmission in Fig. 1;

Fig. 6 is a schematic perspective view of a switching lever fitting of the transmission device in Fig. 1;

Figure 7 is a schematic cross-sectional side view of the retainer fitting of the transmission of Figure 1;

Figure 8 is a schematic cross-sectional side view of the output disc of the transmission of Figure 1;

Fig. 9 is a schematic sectional perspective view of the transmission device in Fig. 1;

Figure 10 is a schematic cross-sectional side view of an alternative embodiment of the axial force generator of the transmission of Figure 1;

Figure 11 is a schematic cross-sectional side view of an alternative embodiment of the transmission of Figure 1;

Figure 12 is a schematic cross-sectional side view of the retainer fitting of the transmission in Figure 11;

Figure 13 is a schematic cross-sectional side view of the optional disengagement means, seen near the axis of the transmission of Figure 11;

Figure 14 is a schematic cross-sectional side view of the optional disengagement means, viewed centrally from the upper and outer portions of the transmission of Figure 11;

Figure 15 is a schematic cross-sectional side view of the axial force generator fitting portion of the transmission of Figure 11;

Fig. 16 is a sectional side view of a variator of the transmission in Fig. 1;

Figure 17 is a schematic cross-sectional side view of an alternative embodiment of the transmission of Figure 1 having two variators;

Figure 18 is a partial cross-sectional view of the transmission, taken from line I-I of Figure 17;

Fig. 19 is a perspective view of the transmission device in Fig. 17;

Figure 20 is a perspective view of the iris plate of the transmission in Figure 17;

Figure 21 is a perspective view of the stator of the transmission in Figure 17;

Figure 22 is a cutaway side view of an optional retainer for the transmission of Figure 17;

Figure 23 is a cross-sectional side view of a ball having the ball groove/leg assembly of Figure 5;

Figure 24 is a cutaway side view of an optional leg of the ball/leg assembly of Figure 5;

Figure 25 schematically depicts the ball and leg assembly and shows the applicable geometrical relationships used to establish the convex curves for the switching guides of the transmissions of Figures 1 and 17;

Figure 26 schematically depicts an obliquely oriented ball and leg assembly and shows the applicable geometrical relationships used to create the convex curves for the switching guides of the transmissions of Figures 1 and 17;

Figure 27 schematically depicts some of the geometrical relationships used to establish the convex curves of the switching guides for the transmissions of Figures 1 and 17;

Figure 28 is a schematic diagram showing the transmission of Figure 1 as a function of a planetary gear set;

Figure 29 is a schematic diagram showing the transmission of Figure 1 showing three planetary gears of a first ratio;

Figure 30 is a schematic diagram showing the transmission of Figure 1 showing three planetary gears of a second ratio;

Figure 31 is a schematic diagram showing the transmission of Figure 1 showing three planetary gears of a third ratio;

Figure 32 is a schematic illustration of the transmission of Figure 1 incorporating planetary gear sets on the output side and in parallel paths;

Figure 33 is a schematic illustration of the transmission of Figure 1 incorporating planetary gear sets on the input side and in parallel paths;

Fig. 34 is a schematic diagram of the transmission of Fig. 1 incorporating a planetary gear set in the output side;

35 is a schematic perspective view of the transmission of FIG. 1 incorporating a planetary gear set in the input side;

Figures 36a, b and c are respectively a cross-sectional side view, a perspective bottom view and a schematic overview of an embodiment of a continuously variable transmission utilizing one torque input and providing two torque output sources;

Figure 37a is a cross-sectional side view of an alternative embodiment of a continuously variable transmission in which the output disc is part of the rotating hub;

Figure 37b is a cross-sectional side view of an alternative embodiment of a continuously variable transmission in which the output disc is part of the stationary hub;

Figure 38 is a side view of an optional ball axle;

Figure 39a is a cross-sectional side view of an optional axial force generator for any of the transmission embodiments described herein;

Figure 39b, c are cross-sectional and perspective views, respectively, of the screw of the optional axial force generator;

Figure 40a is a side elevational view of an alternative linkage mechanism for use with the alternative axial force generator of Figure 39;

Figure 40b is a side elevational view of the alternative linkage mechanism of Figure 40a in an extended configuration.

Detailed ways

Embodiments of the invention will now be described with reference to the corresponding drawings, wherein like reference numerals refer to like elements throughout. The terms used in the specification herein should not be interpreted in a simple or limiting manner, since these terms are used in conjunction with the detailed description of certain specific embodiments of the invention. Furthermore, embodiments of the invention may include several novel features, no single one of which is responsible for its desirable attributes or is essential to practicing the invention described herein.

The transmission described herein is of the type that uses speed adjuster balls with shafts that are inclined as described in US Patent Nos. 6,241,636, 6,322,475 and 6,419,608. The embodiments described in these patents and the embodiments described herein typically have two sides, ie, an input side and an output side, generally separated by a variator section, which will be described below. The driving side of the transmission, that is, the side that receives torque or rotational force into the transmission is called the input side, and the driven side of the transmission, or the side that transmits the torque of the transmission out of the transmission, is called the output side . The input and output discs are in contact with the speed regulator ball. As the speed regulator balls tilt on their axes, the point of rolling contact with one disc moves toward the ball's pole, or axis, with the disc in contact with the ball in a circle of decreasing radius, with the other disc in rolling contact The point of moves towards the ball's equator (equator), so that it makes contact with the ball in a circle of increasing radius. If the axes of the balls are tilted in opposite directions, then the input and output disks respectively contact the balls with the opposite result. In this way, the ratio (or gear ratio) of the rotational speed of the input disc to the rotational speed of the output disc can be varied widely by simply tilting the axis of the speed adjuster ball. The center of the speed regulator ball defines the boundary between the input side and the output side of the transmission, like components located in the input side of the ball and the output side of the ball are generally referred to herein by the same reference numerals. Similar components located on the input and output sides of the transmission will generally have the suffix "a" at the end of their reference designation if they are located on the input side, while components located on the output side of the transmission will generally have the suffix at the end of their respective reference designation "b".

Referring to FIG. 1 , an embodiment of a transmission 100 is depicted having a longitudinal axis 11 about which a plurality of speed adjusting balls 1 are radially distributed. In some embodiments the speed regulating ball 1 is located in angular positions about the longitudinal axis 11, while in other embodiments the speed regulating ball 1 is freely positioned about the longitudinal axis 11. The speed regulating balls 1 are in contact with their input side through the input disc 34 and with their output side through the output disc 101 . The input disc 34 and the output disc 101 are annular discs extending from inner bores near the longitudinal axis on the input and output sides of the speed adjustment ball 1 , respectively, to their respective radial points of contact with the ball 1 . Each of the input disc 34 and the output disc 101 has a contact surface forming a contact area between it and the ball 1 . Typically, as the input disc 34 rotates about the longitudinal axis 11, each part of the contact area of the input disc 34 rotates and contacts each ball 1 sequentially during each rotation. This is similar to output tray 101 . The input disk 34 and output disk 101 may be shaped as simple disks or may be concave, convex, cylindrical or any other shape, depending on the desired input and output configuration. In one embodiment, the input and output disks are formed as spokes to make them lighter in weight sensitive applications. The rolling contact surface of the discs described above that engages the speed adjusting balls may have a flat, concave, convex or other shaped profile depending on the torque and efficiency requirements of the application. The concave profile where the disc meets the ball reduces the axial force needed to prevent slippage, while the convex profile increases efficiency. Furthermore, the balls 1 are in full contact with the idler 18 at their respective radially innermost points. Idler 18 is a generally cylindrical component that rests coaxially about longitudinal axis 11 and helps maintain the radial position of ball 1 . With reference to the longitudinal axis 11 of many embodiments of the transmission, the contact surfaces of the input disc 34 and the output disc 101 can generally be located radially outward from the center position of the ball 1 and the idler 18 is located radially inward from the ball 1 , so that each ball 1 has three contact points with the idler 18 , the input disc 34 and the output disc 101 . In many embodiments, the input disc 34, output disc 101 and idler 18 are rotatable about the same longitudinal axis 11, as will be described in detail below.

Since the embodiments of the transmission 100 described herein are rolling traction transmissions, in some embodiments high axial forces are required to prevent the input disc 34 and output disc 101 from slipping where they are in contact with the balls 1 . Deformation at the contact surfaces of the input disc 34, output disc 101 and idler 18 with the balls 1 is a serious problem as the axial force increases during the transmission of high torques, reducing the efficiency and service life. The amount of torque transmitted through these contact surfaces is limited and is a function of the yield strength of the material from which the ball 1 , input disc 34 , output disc 101 and idler 18 are made. The coefficient of friction of the balls 1, input disc 34, output disc 101 and idler 18 significantly affects the axial force required to transmit a given amount of torque and thus greatly affects the efficiency and service life of the transmission. The coefficient of friction of the rolling elements in a traction drive is a very important variable affecting performance.

Certain coatings may be applied to the surfaces of the balls 1, input disc 34, output disc 101 and idler 18 to improve their performance. In fact, the coatings described above can be beneficially used in the rolling elements of any traction transmission to achieve the same benefits achieved by the transmission embodiments described herein. Certain coatings can beneficially increase the coefficient of friction of the surfaces of the above-mentioned rolling elements. Certain coatings have a large coefficient of friction and also exhibit a variable coefficient of friction that increases with increasing axial force. The large coefficient of friction results in less axial force required for a given torque, thereby increasing the efficiency and service life of the transmission. A variable coefficient of friction increases the maximum torque rating of the transmission by reducing the axial force required to transmit maximum torque.

Certain coatings, such as ceramics and cermets, provide good hardness and wear resistance and can greatly extend the service life of highly loaded rolling elements in rolling traction drives. The ceramic coating (e.g., silicon nitride) has a high coefficient of friction - a variable coefficient of friction that increases with increasing The surface can also increase the service life of these parts when coated with a very thin layer. The thickness of the coating depends on the material used for the coating and can vary depending on the application, but is typically 0.5 microns to 2 microns for ceramics and 0.75 microns to 4 microns for cermets.

The process used to apply the coating described above is important when the balls 1, input disc 34, output disc 101 and idler 18 are made of hardened steel, which is used in many embodiments of the transmissions described herein s material. Some of the processes used to apply ceramics and cermets require high temperatures and will reduce the hardness of the balls 1, input disc 34, output disc 101 and idler 1, compromising performance and causing premature failure. Low temperature application processes are desirable and several are available including low temperature vacuum plasma, DC pulse reactive magnetic sputtering, plasma enhanced chemical vapor deposition (PE-CVD), nonequilibrium magnetophysical vapor deposition, and electroplating. Electroplating treatments are the most attractive due to their low cost and the ability to build custom plating baths to achieve desired coating properties. Immersing the above rolling elements with silicon carbide or silicon nitride in a silicon carbide or silicon nitride electroplating bath with a co-deposited electroless nickel coating or an electrolytic nickel coating is a low temperature solution that works well Suitable for bulky products. It should be noted that other materials than those mentioned above may be used. When applying this treatment, the parts described above are contained in a cage, then dipped into the plating bath and shaken so that the solution contacts all surfaces. The thickness of the coating is controlled by the length of time the above components are submerged in the plating bath. For example, some embodiments will use silicon nitride with a co-deposited electrolytic nickel plating to soak the above assembly for four hours to obtain the proper coating thickness, although this is only one example, many form the coating and control its thickness The methods are well known and can be used to take into account the desired properties, the desired thickness and the substrate or base material from which the above components are made.

Figures 1, 2 and 3 depict an embodiment of a continuously variable transmission 100 that is covered in a housing 40 that protects the transmission 100, contains the lubricant, aligns the components of the transmission 100, and absorbs the transmission 100 force. In some embodiments, shell cap 67 covers shell 40 . Housing cap 67 is generally shaped as a disk with a hole through its center through which the input shaft passes, housing cap 67 having a set of threads at its outer diameter that screw into a corresponding set of threads in the inner diameter of housing 40 . Although in other embodiments the case cap 67 can be secured to the case 40 or held in place by a retaining ring and corresponding groove in the case 40, thus need not be threaded at its outer diameter. In embodiments using fasteners to attach housing cap 67, housing cap 67 extends to the inside diameter of housing 40, thereby enabling housing fasteners (not shown) used to secure housing 40 to the mechanism to which transmission 100 is attached. through the corresponding hole in the housing cap 67. The housing cap 67 of the described embodiment has a cylindrical portion extending from a region around its outer diameter towards the output side of the transmission 100 for additional support of other components of the transmission 100 . At the center of the described embodiment of the transmission 100 there are a plurality of balls 1, typically spherical in shape and distributed radially substantially evenly or symmetrically about the centerline or longitudinal axis 11 of rotation of the transmission 100 around. In the described embodiment, eight balls 1 are used. However, it should be noted that more or fewer balls 1 can be used depending on the use of the transmission 100 . For example, the actuator may comprise 3, 4, 5, 6, 7, 8, 9, 10, 11, 12, 13, 14, 15 or more balls. Preparing more than 3, 4 or 5 balls can spread the force applied to a single ball 1 and the point of contact of a single ball 1 with other components of the transmission 100 more widely, and also can reduce the force that prevents the transmission 100 from being in contact with the ball 100. The force required to slide at the contact surface. Certain embodiments in applications with low torque, high transmission ratios use fewer balls 1 with relatively larger diameters, while certain embodiments in applications with high torque, high transmission ratios may use more ball 1, or a ball 1 with a relatively large diameter. Other embodiments use more balls 1 with relatively small diameters in applications with high torque, low transmission ratios, and where high efficiency is not required. Finally, certain embodiments use fewer relatively small diameter balls 1 in applications that have low torque and do not require high efficiency.

A ball shaft 3 is inserted through a hole through the center of the balls 1 to define a rotational axis of each of the balls 1 . Ball axle 3 is generally an elongated shaft on which ball 1 rotates, and has two ends extending from either side of a bore through ball 1 . Certain embodiments have a cylindrical ball axle 3, although any shape may be used. The ball 1 is mounted for free rotation about a ball axis 3 .

In certain embodiments, bearings (not separately described) are used to reduce the friction between the outer surface of the ball axle 3 and the surface passing through the bore of the corresponding ball 1 . These bearings may be of any type located anywhere along the contact surfaces of the balls 1 and their respective ball shafts 3, and many embodiments will maximize the use of the above bearings by standard mechanical principles well known in dynamic mechanical system design lifespan and use. In some of the embodiments described above, radial bearings are located on each side of the hole through the ball 1 . These bearings can incorporate the inner surface of the bore or the outer surface of the ball shaft 3 as their raceways, or these bearings may comprise separate raceways which fit in the bores of the respective balls 1 and the respective raceways. In a suitable cavity (cavity) formed in the ball shaft 3. In one embodiment, the cavity of the bearing (not shown) is formed by enlarging the holes (at least at both ends) through each ball 1 to an appropriate diameter so that within the cavity thus formed it is secured and held Radial bearings, rollers, balls or other kinds. In another embodiment, the ball shaft 3 is coated with a friction reducing material such as Babbitt, Teflon or other material.

Many embodiments can also minimize the friction between the ball shaft 3 and the ball 1 by using a lubricant in the ball shaft 3 bore. The lubricant is introduced into the bore from a pressure source such as the bore around the ball shaft 3, or through the refling or helical grooves formed on the ball shaft 3 itself. Additional discussions regarding the lubricant of the ball shaft 3 will follow below.

In Figure 1, the axis of rotation of the ball 1 is shown tilted in a direction that places the transmission at a high ratio, where the output speed is greater than the input speed. If the ball shaft 3 is horizontal - parallel to the main axis of the transmission 100, then the ratio of the input rotation rate to the output rotation rate of the transmission 100 is 1:1, where the input and output rotation speeds are equal. In Figure 2, the axis of rotation of the ball 1 is shown inclined in a direction in which the transmission 100 is at a low ratio, meaning that the output speed is less than the input speed. For simplicity, only components that change position and orientation when the transmission 100 shifts are numbered in FIG. 2 .

Figures 1, 2, 4 and 5 illustrate how the axis of the ball 1 can be tilted in operation to switch the transmission 100 . Referring to Fig. 5, a plurality of legs 2 (which in most embodiments are generally struts) are connected to the ball shaft 3 near each end of the ball shaft 3, wherein the ends of the ball shaft 3 extend more than through the ball. 1 at the end of the hole. Each leg 2 extends radially inwardly towards the axis of the transmission 100 from its point of attachment to its respective ball axle 3 . In one embodiment, each of the legs 2 has a through hole receiving a respective end of one of the ball axles 3 . The ball axle 3 preferably extends through the legs 2 such that the ball axle 3 has an exposed end beyond the respective leg 2 . In the described embodiment, the ball shaft 3 advantageously has a roller 4 coaxial and adjustably located at its exposed end. The roller 4 is generally a cylindrical wheel, which is fitted on the ball shaft 3 outside the leg 2 beyond the leg 2 and freely rotates around the ball shaft 3 . The rollers 4 can be attached to the ball axle 3 by spring clips or other mechanisms, or they can be freely attached to the ball axle 3 . The roller 4 may eg be a radial bearing, wherein the outer raceway of the bearing forms a wheel or rolling surface. As shown in Figures 1 and 7, the roller 4 and the ends of the ball shaft 3 fit inside a slot 86 formed by or in a pair of stators 80a, 80b.

One embodiment of a stator 80a, 80b is depicted in FIGS. 5 and 7 . The shown input stator 80a and output stator 80b are arranged generally annularly in the form of parallel discs on either side of the ball 1 about the longitudinal axis 11 of the transmission. The stators 80a, 80b of many embodiments include an input stator disk 81a and an output stator disk 81b, respectively, which are generally annular disks of substantially uniform thickness and having a plurality of apertures as further described below. . Each of the input stator disc 81a and the output stator disc 81b has a first side facing the ball 1 and a second side facing away from the ball 1 . A plurality of stator curved plates 82 are attached to a first side of the stator discs 81a, 81b. The stator curved plate 82 is a curved surface, and the stator curved plate 82 is attached or attached to the stator disks 81a, 81b (each of which has a concave surface 90 facing the ball 1 and a convex surface 91 facing away from the ball 1) and is connected to their respective The stator plate 81 contacts. In some embodiments, the stator curved plate 82 is integrally formed with the stator disks 81a, 81b. The stator curved plates 82 of many embodiments have a substantially uniform thickness and have at least one hole (not shown) for aligning the stator curved plates 82 with each other and attaching the stator curved plates 82 to the stator disc 81 . Many embodiments of the stator curved plate 82 or stator plates 81a, 81b when using a monolithic piece include slots 710 that receive flat spacers 83 that allow the stator curved plate 82 and stator plates 81a, 81b to be further positioned and aligned. arrangement. Flat spacer 83 is a generally flat, and generally rectangular piece of rigid material interconnecting and extending between input stator 80a and output stator 80b. The flat spacers 83 are fixed in slots 710 formed in the stator curved plate 82 . In the described embodiment, the flat spacer 83 is not secured but is attached to the stator curve plate 82, however, in some embodiments the flat spacer 83 is fixed by welding, adhesive or snaps. The object is attached to the stator curved plate 82.

As also shown in FIG. 7 , a plurality of cylindrical spacers 84 (generally cylindrical in shape and having holes at least at each end) are located radially inside the flat spacers 83 and also connect and set the stator disks 81 and stator curve plate 82. The bore of the cylindrical spacer 84 receives a spacer mount 85 at each end thereof. The spacer fixture 85 is designed to clamp and hold together the stator disks 81 a , 81 b , the stator curved plate 82 , the flat spacer 83 and the cylindrical spacer 84 , these assemblies collectively forming a cage 89 . The retainer 89 maintains the radial and angular positions of the balls 1 and aligns the balls 1 with each other.

The axis of rotation of the ball 1 is changed by moving the input-side or output-side leg 2 radially out of the shaft of the transmission 100 , which tilts the ball axis 3 . As this occurs, each roller 4 fits and follows it in a slot 86 slightly larger than the diameter of the roller 4 and formed by the space between each pair of adjacent curved stator plates 82 . Thus, the rollers 4 roll along the surfaces of the sides 92, 93 of the stator curved plates 82 (the first side 92 and the second side 93 for each stator curved plate 82) to keep the plane of movement of the ball shaft 3 in line with the transmission. The longitudinal axis 11 of 100 coincides. In many embodiments, each roller 4 rolls in the first side 92 of the stator curve plate 82 on the input side of the transmission 100 and on the first side 92 of the corresponding output stator curve plate 82 . In the exemplary embodiment described above, the force of the transmission 100 prevents the roller 4 from contacting the second side 93 of the stator curved plate 82 in normal operation. The diameter of the rollers 4 is slightly smaller than the width of the slots 86 formed between the stator curved plates 82, which creates a small gap between the edges of the slots 86 and the circumference of each respective roller. If the opposing sets of stator curve plates 82 on the input stator 80a and output stator 80b are perfectly aligned, the clearance between the periphery of the roller 4 and the slot 86 will allow the ball axle to tilt slightly and align with the longitudinal axis of the transmission 100. 11 is not aligned. This situation creates slippage, a condition that causes the ball axle 3 to move slightly laterally, which reduces the overall shifting efficiency. In some embodiments, the stator curved plates 82 on the input and output sides of the transmission 100 may be slightly offset from each other so that the ball shaft 3 remains parallel to the axis of the transmission 100 . Any tangential forces, mainly transaxial forces, that the ball 1 may exert on the ball shaft 3 are absorbed by the ball shaft 3 , the roller 4 and the first side 92 , 93 of the stator curved plate 82 . As the transmission 100 is switched to a lower or higher gear ratio by changing the axis of rotation of the ball 1, each of a pair of rollers 4 located at opposite ends of a single ball axis 3 passes in opposite directions by winding and Roll down each side of the groove 86 to move along their respective corresponding groove 86 .

Referring to FIGS. 1 and 7 , the retainer 89 can be rigidly connected to the shell 40 using one or more shell connectors 160 . Shell connector 160 extends generally vertically from the radially outermost portion of flat spacer 83 . The shell connector 160 can be fixed to the flat spacer 83 or integrally formed with the flat spacer 83 . The outer diameter roughly formed by the exterior of the housing connector 160 is substantially the same size as the inner diameter of the housing 40, and the holes in the housing 40 and the housing connector 160 are standard or dedicated to rigidly connect the housing connector 160 to the housing 40. Fixtures are used to stabilize and prevent the shell 40 from moving. The shell 40 has mounting holes for attaching the shell 40 to a frame or other structure. In other embodiments, the shell connector 160 can be formed as part of the shell 40 and provide a location for attaching a flat spacer 83 or other component of the retainer 89 to mobilize the retainer 89 .

FIGS. 1 , 5 and 7 depict an embodiment comprising a pair of stator wheels 30 attached to each leg 2 , the stator wheels 30 rolling on the concave surface 90 of the curved plate 82 along a path near the edges of the sides 92 , 93 . The stator wheel 30 is usually attached to the leg 2 in the region where the ball axle 3 passes through the leg 2 . The stator wheel 30 can be attached to the leg 2 with a stator wheel pin 31 penetrating through a hole in the leg 2 (typically perpendicular to the ball axis 3 ) or by other attachment methods. The stator wheel 30 is coaxially and slidably mounted to the stator wheel pin 31 and secured with standard fasteners such as snap rings. In some embodiments, the stator wheel 30 is a radial bearing having an inner raceway mounted to the stator wheel pin 31 and an outer raceway forming a rolling surface. In some embodiments, a stator wheel 30 is provided on one side of the leg 2 with sufficient spacing from the leg 2 to allow the stator wheel 30 to move along the concave surface 90 relative to the transmission 100 when the transmission 100 is switched. The longitudinal axis 11 of the device 100 rolls radially. In certain embodiments, the concave surfaces 90 are shaped such that they are concentric around the extent of the radius formed by the center of the ball 1 to the longitudinal axis 11 of the transmission 100 .

Still referring to FIGS. 1 , 5 , 7 , there is depicted a guide wheel 21 attachable to the end of the leg 2 closest to the longitudinal axis 11 of the transmission 100 . In the described embodiment, the guide wheels 21 are inserted into grooves formed in the ends of the legs 2 . The guide wheels 21 are held in place in the slots of the legs 21 using guide wheel pins 22 or by other arbitrary attachment methods. The guide wheels 21 are coaxially and slidably mounted above guide wheel pins 22 which are inserted into holes formed in the legs 2 on each side of the guide wheels 21 and perpendicular to the plane of the slot. In some embodiments, the legs 2 are designed to deflect elastically relatively slightly, in order to allow for tolerances in the manufacture of the components of the transmission 100 . Ball 1 , leg 2 , ball axle 3 , roller 4 , stator wheel 30 , stator wheel pin 31 , guide wheel 21 and guide wheel pin 22 collectively form a ball/leg assembly 403 , see FIG. 5 .

Referring to the embodiment depicted in FIGS. 4 , 6 and 7 , switching is activated by rotating the lever 10 located outside the housing 40 . Rod 10 is used to roll up and unwind flexible input cable 155a and flexible output cable 155b, flexible input cable 155a and flexible output cable 155b are connected to rod 10 at their respective first ends, and with mutually opposite The direction is wound on the rod 10. In some embodiments, the input cable 155a is wound around the rod 10 in a counterclockwise direction and the output cable 155b is wound around the rod 10 in a clockwise direction (the rod 10 is viewed from right to left in FIG. 6 ). The input cable 155a and the output cable 155b extend through the aperture in the housing 40 and then through the input flexible cable sleeve 151a and the first end of the output flexible cable sleeve 151b. The input flexible cable sheath 151a and the output flexible cable sheath 151b in this depicted embodiment are flexible elongated tubes that guide the input cable 155a and the output cable 155b radially inwardly towards the longitudinal axis 11, It is then threaded longitudinally out of the holes in the stator disks 81a, 81b and again radially inwards to where the input and output flexible cable sleeves 151a, 151b are embedded and attached to the input and output rigid cable sleeves 153a respectively , 153b. The input and output rigid cable sleeves 153a, 153b are inflexible tubes through which the cables 155a, 155b pass and are guided radially inwardly from the second ends of the input and output flexible cable sleeves 151a, 151b and then The cables 155a, 155b are guided longitudinally through the bores of the stator disks 81a, 81b and towards the second end of the rigid cable sheath 153a, 153b near the idler 18 . In many embodiments, cables 155a, 155b are attached at their second ends to input switching guide 13a and output switching guide 13b using conventional cable fixtures or other suitable attachment means (further below. describe). As will be described further below, the switching guides 13a, 13b determine the position of the idler 18 axially along the longitudinal axis 11 and radially of the legs 3, thus changing the shaft and transmission of the ball 1 100 gear ratio.

If the user rotates the pole 10 counterclockwise (right to left with respect to the axis of the pole 10, as shown in FIG. The output cable 155b is wound to the pole 10 . Thus, the second end of the output cable 155b applies tension to the output switching guide 13b, and the input cable 155a is released from the rod 10 by a corresponding amount, which pushes the idler 18 axially. Move toward the output side of the transmission 100 and shift the transmission 100 toward low.

Still referring to Figures 4, 5 and 7, the described switching guide means 13a, 13b are each in the form of a ring having an inner diameter and an outer diameter, and are shaped to have two sides. The first side is generally a straight surface that dynamically contacts and coaxially supports the idler 18 via two sets of idler bearings 17a, 17b associated with respective switching guides 13a, 13b. The second side (the side facing away from the idler 18) of each switching guide 13a, 13b is the cam side, which goes from the straight or flat radial surface 14 towards the inner diameter of the guide 13a, 13b towards the guide 13a. , 13b of the outer diameter of the convex curve plate 97 transformation (transition). At the inner diameter of the guide means 13a, 13b, longitudinal sleeves 417a, 417b extend axially from these guide means 13a, 13b towards the opposite guide means 13a, 13b to cooperate with the sleeves 417a, 417b. In some embodiments, as shown in FIG. 4, the sleeve of the input side switching guide 13a has a portion of its inner diameter drilled to receive the sleeve of the output switching guide 13b. Accordingly, a portion of the outer diameter of the socket of the output switching guide 13b is removed to allow a portion of the socket 417a, 417b to be inserted into the socket 417a, 417b of the input switching guide 13a. This provides additional stability to the switching guide means 13a, 13b of the above-described embodiments.

The cross-sectional view of the switching guide 13a, 13b shown in FIG. The longitudinal axis 11, up to the radial point where the guide wheel 21 comes into contact with the switching guide means 13a, 13b. From this point the convexly shaped curved profile of the switching guide 13a, 13b protrudes towards the periphery of the switching guide 13a, 13b. In some embodiments, the convexly curved plates 97 of the switching guides 13a, 13b are not radial, but consist of multiple radii, or have a hyperbolic, asymptotic, or other shape. As the transmission 100 shifts low, the input guide wheel 21a rolls toward the longitudinal axis 11 on the flat portion 14 of the shift guide 13a and the output guide wheel 21b rolls away from the longitudinal axis 11 on the convex curve portion 97 of the shift guide 13b. The input guide wheels 21a, 21b can be obtained by screwing the male threaded socket of the input switching guide 13a to the female threaded socket of the output switching guide 13b (and vice versa), and threading the switching guides 13a, 13b joined together and connected to each other. One switching guide 13a, 13b (input or output) can also be pushed into the other switching guide 13a, 13b. The switching guide devices 13a, 13b can also be joined together by other methods, such as glue, metal adhesive, welding or other methods.

The convex curved plates 97 (as cam surfaces) of the two switching guides 13a, 13b are in contact with and pushed against the plurality of guide wheels 21, respectively. The flat surface 14 and the convex curve 97 of each switching guide 13a, 13b are in contact with the guide wheel 21 so that when the switching guide 13a, 13b is moved axially along the longitudinal axis 11, the guide wheel 21 moves in a generally radial direction Ride along the surfaces 14, 97 of the switching guides 13a, 13b, forcing the legs 2 radially towards or away from the longitudinal axis 11, thereby changing the angle of the ball axis 3 and the associated axis of rotation of the ball 1 .

Referring to FIGS. 4 and 7 , the idler wheel 18 of some embodiments is located in a slot formed between the first side and the sleeve portion of the shift guides 13a, 13b and thus moves in unison with the shift guides 13a, 13b. In certain embodiments, idler 18 is generally tubular having an outer diameter and a central portion along its inner diameter that is substantially cylindrical and has input and output idler bearings 17a at each end of its inner diameter, 17b. In other embodiments, the outer and inner diameters of the cam 18 can be non-uniform and can vary or have arbitrary shapes, such as sloped or curved. The idler 18 has two sides, one close to the input stator 80a and the other close to the output stator 80b. The idler bearings 17a, 17b provide rolling contact between the idler 18 and the switching guides 13a, 13b. Idler bearings 17 a , 17 b are located coaxially around the sleeve portion of the shift guides 13 a , 13 b , allowing the idler 18 to rotate freely about the axis of the transmission 100 . A sleeve 19 is arranged around the longitudinal axis 11 of the transmission 100 and inside the inner diameter of said switching guides 13a, 13b. The sleeve 19 is generally a tubular component and is held in operative contact with the inner bearing race surface of each of the shift guides 13a, 13b via an input sleeve bearing 172a and an output sleeve bearing 172b. The sleeve bearings 172a, 172b rotate the sleeve 19 by rolling along the outer bearing races complementary to the races of the switching guides 13a, 13b. Idler 18 , idler bearings 17 a , 17 b , sleeve 19 , shift guides 13 a , 13 b and sleeve bearings 172 a , 172 b together form idler assembly 402 , as shown in FIG. 4 .

Referring to FIGS. 4 , 7 and 8 , the inner diameter of the sleeve 19 of some embodiments is threaded to receive the threaded insertion of an idler rod 171 . The follower rod 171 is a generally cylindrical rod and is arranged along the longitudinal axis 11 of the transmission 100 . In some embodiments, follower rod 171 is threaded at least partially along its length to allow insertion of sleeve 19 . The first end of the follower rod 171 (facing the output side of the transmission 100) is preferably threaded through the sleeve 19 and extends through the output side of the sleeve 19 where the first end of the follower rod 171 is inserted into the output disc 101 inside diameter.

As shown in FIG. 8 , in some embodiments output disc 101 is a generally conical disc formed as spokes to reduce weight and has a tubular sleeve portion extending axially from its inner diameter toward the outside of transmission 100 . The output disc 101 transmits output torque to a drive shaft, wheel or other mechanical device. The output discs 101 are in contact with the balls 1 on their output side and rotate at a different speed (in a ratio other than 1:1) from the input rotation of the transmission. The output disc 101 serves to guide and center the follower rod 171 at its first end so that the sleeve 19 , idler 18 and switching guides 13 a , 13 b are concentric with the axis of the transmission 100 . As an option, an annular bearing may be located above follower rod 171 (between follower rod 171 and the inner diameter of output disc 101 ) to reduce friction. The follower rod 171, sleeve 19, shift guides 13a, 13b and idler wheel 18 are operatively connected and all cooperatively move axially when the transmission 100 is shifted.

Referring to FIG. 2, the conical spring 133 (located between the input shift guide 13a and the stator 80a) shifts the transmission 100 low. Referring to FIG. 1 , the output disc bearing 102 (in contact with the bearing race near the periphery of the output disc 101 ) absorbs axial forces generated by the transmission 100 and transmits them to the housing 40 . The shell 40 has a corresponding bearing race to guide the output disc bearing 102 .

Referring to Figures 4, 5 and 7, the limitation of the axial movement of the switching guides 13a, 13b defines the switching range of the transmission. Axial movement is limited by inner faces 88a, 88b in the stator discs 81a, 82b, which are in contact with the switching guides 13a, 13b. The switching guide 13a is in contact with the inner face 88a in the input stator disk 81a at a very high transmission ratio and the switching guide 13b is in contact with the inner face 88b in the output stator disk 81b at a very low transmission ratio. In many embodiments, the curvature function of the convex curve 97 of the switching guide 13a, 13b depends on the distance from the center of the ball 1 to the center of the guide wheel 21, the radius of the guide wheel 21, the distance between the two guide wheels 21 and the ball 1. The angle between the lines formed between the centers of , and the angle at which the axis of the ball 1 is inclined. Examples of the above relationship will be described below with reference to FIGS. 25 , 26 and 27 .

Referring now to the embodiment described in FIGS. 1 , 5 and 7 , one or more stator wheels 30 can be attached to each leg 2 using stator wheel pins 31 inserted through holes in the respective leg 2 . The stator wheel pin 31 is sized appropriately and allows the stator wheel 30 to rotate freely thereon. The stator wheel 30 rolls along the concave curved surface 90 of the stator curved plate 82 facing the ball 1 . The stator wheel 30 provides axial support to prevent the legs 2 from moving axially and ensures that the ball shaft 3 can be easily tilted when the transmission 100 is shifted.

Referring to FIGS. 1 and 7 , the input disc 34 formed as a spoke (located adjacent to the stator 80a ) partially encapsulates the stator 80a but does not contact it. The input disk 34 may have two or more spokes, or may be a solid disk. The spokes reduce weight and facilitate assembly of transmission 100 . In other embodiments, solid disks may be used. The input disc 34 has two sides, a first side in contact with the ball 1 and a second side facing away from the first side. The input disc 34 is generally annular and coaxially fits on its inner diameter a set of female threads or nuts 37 and extends radially therefrom. If the type of shell 40 used encloses the ball 1 and input disc 34 and is mounted to a rigid support structure 116 (e.g., a chassis or frame with conventional bolts inserted through bolt holes in the flange of the shell 40), the input disc The outer diameter of 34 is designed to fit within shell 40 . As mentioned above, the input disc 34 is in rotational contact with the ball 1 along the circumferential sliding table or bearing surface on the lip of its first side (the side facing the ball 1 ). As also mentioned above, some embodiments of the input disc 34 have a set of female threads 37 inserted into its inner diameter or nuts 37 which are threaded onto the screw 35 to engage the input disc 34 and screw 35 .

Referring to FIGS. 1 and 4 , the screw 35 is attached to and rotated by a drive shaft 69 . The drive shaft 69 is generally cylindrical and has an inner bore, a first end axially facing the output side, a second end axially facing the input side, and a generally constant diameter. At a first end, a drive shaft 69 is rigidly attached to and rotated by an input torque device (typically a gear, sprocket or crankshaft from an electric motor). The drive shaft 69 has axial splines 109 extending from the second end of the drive shaft 69 to engage and rotate a corresponding set of splines formed on the inner diameter of the screw 35 . A set of central drive shaft ramps 99 (on a first side generally a set of raised ramped surfaces in an annular disc coaxial with drive shaft 69) have mating prongs that mate with splines 109 in drive shaft 99 , is rotated by the drive shaft 69 and can move axially along the drive shaft 69 . The pin ring 195 contacts the second side of the central drive shaft ramp 99 . The pin ring 195 is a rigid ring coaxially on the follower rod 171 and is axially movable and has a transverse hole which functions to keep the idler pin 196 in alignment with the follower rod 171 . The idler pin 196 is an elongated rigid rod slightly longer than the diameter of the pin ring 195, and is inserted through the elongated slot 173 in the follower rod 171, and when it is inserted into the hole of the pin ring 195 at its first position. The pin ring 195 extends slightly beyond the one and second ends. The elongated slot 173 in the follower rod 171 allows the follower rod 171 to move to the right (when viewed in FIG. 1 ) without contacting the pin 196 when the transmission 100 is shifted from a 1:1 ratio to high. However, when the transmission 100 shifts from a 1:1 ratio to low, the side of the input end of the elongated slot 173 contacts the pin 196 and then operatively contacts the central drive shaft ramp 99 through the pin ring 195 . Thus, when the transmission is in gear ratios of 1:1 and lower, the follower rod 171 is operatively connected to the central drive shaft ramp 99 so that when the follower rod 171 moves axially, the central drive shaft ramp 99 Move together with follower rod 171. The ramped surfaces of the central drive shaft ramps 99 may be helical, curved, linear or any other shape and operably contact a set of corresponding central bearing disc ramps 98 . The central bearing disc ramp 98 has a complementary and opposite ramp to the central drive shaft ramp 99 . On a first side (which faces the output side of the transmission 100 ), the central bearing disc ramp 98 faces the central drive shaft ramp 99 and contacts and is driven by the central drive shaft ramp 99 .

The central bearing disc ramp 98 is rigidly attached to the bearing disc 60 , which is generally an annular disc arranged to rotate axially about the longitudinal axis 11 of the transmission 100 . The bearing disk 60 has a bearing race near the periphery of its side facing away from the ball 1 , which is in contact with the bearing disk bearing 66 . Bearing disk bearing 66 is an annular thrust bearing located at the periphery of bearing disk 60 and between bearing disk 60 and input disk 34 . Bearing disc bearings 66 provide axial and radial support for bearing disc 60 and are in turn supported by bearing races in housing cap 67 which cooperates with housing 40 to partially enclose the interior of transmission 100 .

Referring to FIG. 1 , housing cap 67 is a generally annular disk extending from drive shaft 69 and having a tubular portion extending from (or near) its periphery toward an outer end and having a bore through its center. The housing cap 67 absorbs axial and radial forces generated by the transmission 100 and seals the transmission 100 to prevent lubricant leakage and ingress of contaminants. Shell cap 67 is fixed in some embodiments and is rigidly attached to shell 40 using conventional fastening methods, or has male threads on its outer diameter that mate with corresponding female threads in the inner diameter of shell 40 . As mentioned above, the housing cap 67 has a bearing race in contact with the bearing disc bearing 66 near the outer periphery of the bearing disc 60 (inboard of the outer end of the tubular extension of the housing cap 67). The housing cap 67 also has a second bearing race facing the outside near the inner diameter of its annular portion, the second bearing race cooperating with the drive bearing 104 . Drive bearing 104 is a combined thrust and radial bearing that provides axial and radial support for drive shaft 69 . The drive shaft 67 has a bearing race formed on its outer diameter facing the input side that cooperates with the drive bearing 104 and transmits the axial force generated by the screw 35 to the housing cap 67 . Input bearing 105 adds support to drive shaft 69 . The input bearing 105 is located axially above the drive shaft 69 and cooperates with a third race on the inner diameter of the housing cap 67 facing the input side of the transmission 100 . A conical nut 106 (typically a cylindrical nut with a running surface designed to provide a running surface for the input bearing 105 ) is threaded onto the drive shaft 69 and supports the input bearing 105 .

Referring to the embodiment depicted in FIG. 1 , a set of a plurality of peripheral ramps 61 , generally forming a ring about the longitudinal axis 11 , is rigidly attached to the bearing disc 60 . The peripheral bevels 61 are a plurality of inclined surfaces radially surrounding the longitudinal axis 11 and facing the outside, which abut against or are formed in the bearing disk 60 . The sloped surfaces may be curved, helical, linear or other shapes, and each of which establishes a wedge that produces an axial force applied to one of the plurality of ramp bearings 62 . A ramp bearing 62 , which is spherical but could also be cylindrical, conical, or other geometric shape, is provided in a bearing holder 63 . The bearing retainer 63 of the embodiment described here is generally annular and has a plurality of holes that receive a single ramp bearing 62 . A set of input disc ramps 64 is rigidly connected to or formed as part of the input disc 34 . The input disc ramp 64 in some embodiments is complementary to the peripheral ramp 62 having a ramp facing the input side. In other embodiments, the input disc ramp 64 is in the form of a bearing race that rigidly aligns and centers the ramp bearing 62 . The ramp bearing 62 responds to changes in torque by rolling up or down the ramped surfaces of the peripheral ramp 61 and input disc ramp 64 .

Referring now to FIGS. 1 and 4 , the axial force generator 160 is made up of a number of components that generate and apply an axial force to the input disc 34 to increase the normal contact force between the input disc 34 and the ball 1, The axial force generator 160 is a friction component for the input disc 34 to be used in the rolling ball 1 . The transmission 100 generates sufficient axial force such that the input disc 34, ball 1 and output disc 101 do not slip, or only slip by an acceptable amount, at their contact points. As the magnitude of torque applied to the transmission 100 increases, appropriate additional axial force is required to prevent slippage. Additionally, more axial force is required to prevent slippage at rates below, above, or equal to 1:1. However, providing too much axial force above or equal to 1:1 will shorten the life cycle of the transmission 100 and reduce efficiency and/or require larger components to absorb the increased axial force. Ideally, the axial force generator 160 varies the axial force applied to the ball 1 as the transmission 100 is switched and as the torque changes. In some embodiments, transmission 100 achieves both of the above goals. The screw 35 is designed and configured to provide an axial force that is separate from the axial force generated by the peripheral ramp 61 . In some embodiments, the screw 35 generates less axial force than the peripheral ramp 61 , although in other transmission 100 variations, the screw 35 is configured to generate a greater axial force than the peripheral ramp 61 . After the torque is increased, the screw 35 rotates and enters the nut 37 slightly to increase the axial force in proportion to the increase in torque. If the transmission 100 is in a 1:1 ratio and the user or vehicle shifts to a low speed, the follower rod 171 moves axially towards the input side along with sleeve 19, sleeve bearing 172, shift guides 13a, 13b and idler 18 . The follower rod 171 is in contact with the central drive shaft ramp 99 through the pin 196 and the pin ring 195 , which moves the central drive shaft ramp 99 axially toward the output side. The ramped surface of central drive shaft ramp 99 contacts the opposing ramped surface of central bearing disc ramp 98 , which causes central bearing disc ramp 98 to rotate bearing disc 67 and engage peripheral ramp 61 with ramp bearing 62 and input disc ramp 64 . The central drive shaft ramp 99 and the central bearing disc ramp 98 perform the function of torque splitting, ie, shifting some of the torque from the screw 35 to the peripheral ramp 61 . This increases the percentage of transmitted torque directed through the peripheral bevel 61 and, since the peripheral bevel 61 is torque sensitive (as described above), the axial force generated increases.

Still referring to FIGS. 1 and 4 , upon switching to low, the idler 18 moves axially towards the output side and is pulled low by the reaction force in the interface. The more the idler wheel 18 moves low, the stronger it is pulled. This "idler pull" (increasing with normal force across the interface and switching angle) also occurs when switching to high. The collection of shear forces in the interface causes idler pull to occur, the effect of which is called spin. Spin occurs in three contact surfaces, the points where the ball contacts the input disc 34 , the output disc 101 and the idler 18 . The resultant force generated by the spin at the contact between the idler 18 and the ball 1 is minimal compared to the resultant force generated by the spin between the ball 1 and the input and output discs 34,101. Since the minimum spin is generated at the contact surface where the idler 18 and ball 1 interface, this contact surface is ignored for the following explanation. Spin can be seen as a loss of efficiency in the interface between the input disk 34 and ball 1 and the output disk 101 and ball 1 . The shear force generated by the spin is perpendicular to the direction of rotation of the ball 1 and disk 34,101. At a 1:1 ratio, the shear forces due to spin or contact spin are equal and opposite at the input and output contact surfaces, so are essentially cancelled. In this case there is no axial pull on the idler 18 . However, when the transmission 100 is shifted towards low, for example, the interface at the input disc 34 and the ball 1 moves further away from the axis and pole of the ball 1 . This reduces the spin and the resulting shear perpendicular to the direction of rotation. At the same time, the output disk 101 and ball 1 interface moves closer to the axis and pole of the ball 1, which increases the spin and resultant shear. This creates a situation where the shear forces generated by the spin are unequal on the input and output sides of the transmission 100, and because the shear force in the output contact is greater, the contact surface between the output disc 101 and the ball 1 moves closer to Axis of ball 1. The lower the transmission 100 is switched, the stronger the shear force exerted on the ball 1 in the contact surface. On switching to high, the shear generated by the spin on ball 1 exerts a force in the opposite direction. The leg 2 attached to the ball shaft 3 applies this pulling force to the switching guides 13a, 13b, since the switching guides 13a, 13b are operatively attached to the idler 18 and the sleeve 19, the axial force is transmitted to the following Moving bar 171. As the mana over touch increases, the impact of the touch spin at all ratios increases and efficiency decreases.

Still referring to FIGS. 1 and 4, when the transmission 100 is switched to low, the pulling force transmitted to the follower rod 171 causes an axial force toward the left, as shown in FIG. Switch to peripheral bevel 61 . When the transmission 100 is shifted lower, the pulling force of the follower rod 171 is greater, causing relative movement between the central drive shaft slope 99 and the central bearing disk slope 98, and switching to the outer peripheral slope 61 for greater torque. This reduces the torque transmitted through the screw 35 and increases the torque transmitted through the peripheral bevel 61, resulting in an increase in the axial force.

The decoupling device (consisting of several parts that will be described) is described with reference to FIGS. 1 and 9 . The disengagement device is located between the input disc 35 and the bearing disc 60 and releases the transmission 100 when the output rotation is greater than the input rotation. The breakaway device is made up of several parts, including an input disc connector 121 (typically a cylindrical elongated pin rigidly attached at its periphery to the input disc 34) which runs from the input disc 35 towards the bearing disc 60 in substantially parallel The direction of the longitudinal axis 11 of the transmission 100 protrudes. The input disc connector 121 engages the clutch lever 122 at a first end. The clutch lever 122 is generally an L-shaped flat piece of rigid material and has a first end extending as its short leg and a second end extending as its long leg, pivoting from a joint at the intersection of its legs. Connected to the preloader (preloader) 123. The engagement of the input disc connector 121 with the first end of the trip lever 122 is a sliding engagement and allows relative movement between the connector 121 and the trip lever 122 . The coupling of the trip lever 122 is formed through a through hole on the preloader 123 . The preloader 123 is a flexible, elongated rod, which may also be square, flat, or other cross-sectional shape, and is attached at its first end to a bore extending radially through the bearing holder 63 and at its second end. The two ends are then rigidly attached to the drive shaft 69 . The pre-loader 123 can bias the ramp bearing 62 upward to the outer peripheral ramp 61, the pre-loader 123 can also disengage the input disk 34 from the ball 1 during the time the disengagement device is activated, and can be used as a tool for other components such as , detachment device 120 components) attachment means. A pawl 124 is attached to the clutch lever 122 . The pawl 124 is generally wedge-shaped, tapering to a point at its first end, and circular at its second end with a through hole. The pawl pin 125 is inserted into a hole at the second end of the trip lever 122 thereby attaching the pawl 124 to the trip lever 122 and allowing the pawl 124 to rotate about the pawl pin 125 . The pawl 124 cooperates with and contacts a disc ratchet 126 which has teeth around its circumference and lies flat against the rear of the clutch lever 122 . In the center of the ratchet 126 there is a hole through which the preloader 123 adjoins the clutch lever 122 , which hole is also radially inwards towards the longitudinal axis 11 of the transmission 100 . Ratchet 126 is held in place by conventional mounts and is able to rotate about preloader 123 . A ratchet bevel 127 (a gear having beveled teeth on its periphery) is rigidly and coaxially attached to and forms part of the ratchet 126 . The teeth in the ratchet bevel 127 mesh with the bevel gear 128 . Bevel gear 128 is a ring rigidly attached to bearing disk 60 in this described embodiment, however, it may be attached to other rotating components such as drive shaft 69 and central drive shaft ramp 99 . The bevel gear 128 has teeth around its periphery that mate with the teeth of the ratchet 126 . A main spring 129 (coil spring having multiple turns, as shown in FIG. 11 ) is coaxially disposed about the longitudinal axis 11 of the transmission 100 and is attached at its first end to the input disc 34 and at its second end to Connected to the bearing plate 60. The main spring 129 presses the input disc 34 to rotate about or loosen from the screw 35 so that the input disc 34 comes into contact with the ball 1 .

Still referring to Figures 1 and 9, the ball 1 is driven by the output disc 101 when the input rotation of the transmission 100 ceases and the output disc 101 continues to rotate by one or more wheels, transmission or other output rotating means. The balls 1 are then rotated into the input disk 34 in a first direction to "wrap around" the screw 35 and separate from the balls 1 . Input disc connector 121 , rotated in the same first direction by input disc 34 , contacts clutch lever 122 and pawl 124 and causes them to rotate in the first direction. Pawl 124 is biased into contact with the teeth of ratchet 126 by a pawl tensioner (not shown), which may be a torsion spring positioned coaxially over pawl pin 125 . As the pawl 124 passes over the teeth of the ratchet 126, the pawl 124 locks on the teeth of the ratchet 126, preventing the input disc 34 from being released from the screw 35 in the second direction and coming into contact with the ball 1 again, like The bias of the main spring 129 produces the same result. Since the ratchet bevel 127 (which is part of the ratchet 126 ) has teeth that interlock with the bevel gear 128 (which does not rotate), the ratchet 126 is prevented from rotating in the second direction.

When the input rotation of transmission 100 is restarted, ratchet bevel 127 is rotated in a first direction by bearing plate 60, which causes ratchet bevel 127 and ratchet 126 to rotate in a second direction, thereby causing pawl 124 to rotate in a second direction. Rotation in the second direction enables the main spring 129 to bias the input disc 34 to unwind from the screw 35 and come into contact with the ball 1 in the second direction. It is important to note that bearing holder 63 attached at a first end to pre-loader 123 rotates pre-loader 123 relative to input disc 34 as input disc 34 rotates in a first direction. This is because the ramp bearing 62 rotates relative to the input disc 34 as the input disc 34 rotates in the first direction. Likewise, when the input rotation of the transmission 100 is restarted, the bearing disc 60 rotates relative to the preloader 123 due to the same relative rotation. These actions engage and disengage the disengagement device 120 .

Referring to Figures 1 and 15, a latch 115 is rigidly attached to the side of the output disc 34 facing the bearing disc 60, and is rigidly attached to the first end of the hook lever (hook lever) 113 at both ends. The hook 114 engages. The hook lever 113 is an elongated post having a hook 114 at its first end and a hook hinge 116 at its second end. The locking means 115 has an engagement area or opening which is larger than the width of the hook 114 and which provides radial movement of the hook 114 within the boundaries of the hook 114 relative to the longitudinal axis 11 when the input disc 34 and the bearing disc 60 are moved relative to each other. additional space. Hook hinge 116 engages intermediate hinge 119 and forms a hinged joint with first hinge pin 111 . Intermediate hinge 119 is integrally formed with input tray lever 112, which is generally an elongated post with two ends. At a second end of the input tray lever 112 there is an input tray hinge 117 which engages the hinge bracket 110 using a second hinge pin 118 . Hinge bracket 110 is generally support hook 114, hook lever 113, hook hinge 116, first hinge pin 111, middle hinge 119, input disc lever 112, second hinge pin 118 and input disc hinge 117 and is rigidly attached to the bearing disc 34 on the side facing the input disc 34 . When the latch 115 and hook 114 are engaged, the ramp bearing 62 is prevented from rolling into areas on the peripheral ramp 61 that do not provide the correct amount of axial force to the drive plate 34 . This positive engagement ensures that all rotational forces applied to the ramp bearing 62 through the peripheral ramp 61 are transmitted to the input disc 34 . The preloader 123 is attached to the drive shaft 69 at a first end and extends radially outward. The preloader 123 contacts the input disc lever 112 at a second end to bias the input disc 34 away from the ball 1 such that the input disc 34 is biased to remain unattached when the input disc 34 is separated from the ball 1 .

Referring to FIG. 10 , a cross-sectional view of an alternative axial force generator of transmission 100 is disclosed. For simplicity, only the differences between the axial force generator described above and the axial force generator in Figure 10 are shown. The axial force generator shown includes one or more return rods 261 . The return levers 261 are generally flat, irregularly shaped cam members and each have an eccentrically positioned pivot hole with a first side radially inward of the pivot hole and a second side radially outward of the pivot hole. The first side of each reversing rod 261 fits into an elongated slot 173 in the follower rod 171 . When the transmission 200 is shifted towards low, the end of the elongated slot 173 contacts the first side of the reverse lever 261 formed by the reverse pin 262 inserted into the pivot hole of the reverse lever 261 Axis rotation. At the first end where the end of the elongated slot 173 contacts, the first side of each reversing lever 261 moves towards the output side of the transmission 100, and the second side of the reversing lever 261 moves towards the input side of the transmission 100, so that This performs the camming function of the reverse lever 261 . By increasing and decreasing the lengths of the first and second sides, the return levers 261 can be designed to reduce the distance they travel axially towards the input side, and to increase the force they generate. The return levers 261 can be designed in such a way as to create the mechanical benefit of adjusting the axial force they generate. On their second side, each return rod 261 is in contact with the output side of the central screw ramp 298 when the transmission 100 is shifted towards low. Each return lever 261 is attached to a control lever ring 263 by a return pin 262 which can be pressed or screwed into a hole in the control lever ring 263 and holds the return lever 261 in place . Lever ring 263 is an annular shaped device that fits around and slides axially along follower rod 171 and has one or more rectangular slots through which the return rod 261 is inserted and positioned.

Still referring to the embodiment depicted in FIG. 10 , a set of central screw ramps 299 are rigidly attached to and rotatable by screw 35 . The central screw ramp 299 of this embodiment is similar to the central screw ramp 99 described in FIG. Bevels on both sides. As the transmission 100 shifts toward low, the second side of the reverse rod 261 pushes against the first side of the central screw ramp 299 . The central screw ramp 299, which is splinedly engaged to the drive shaft 69 via the aforementioned spline 109, is rotated by the drive shaft 69, is axially movable along the longitudinal axis 11, and is facing the input side of the transmission 100 except that the central screw ramp 299 is Other than the output side, it is similar to the central drive shaft ramp 99 in the previous embodiment. The central screw ramps 299 contact an opposing set of central bearing disc ramps 298 which are free to rotate relative to the drive shaft 69 and are similar to the central bearing disc ramps 98 described in FIG. is facing the output side of the transmission 100 rather than the input side). As the return rod 261 pushes the central screw ramp 299 axially toward the central bearing disc ramp 298, a relative rotation of the ramps of the central screw ramp 299 and the central bearing disc ramp 298 occurs, which rotates the bearing disc 60 so that the peripheral ramp 61 engages, thereby switching the torque to the peripheral ramp 61 and increasing the amount of axial force generated.

Referring now to FIG. 11 , a cross-sectional view of an alternative embodiment of the transmission 100 of FIG. 1 is disclosed. For the sake of simplicity, only the differences between the preceding transmission 100 and this transmission 300 are described. The transmission 300 has an alternate retainer 389, an alternate disengagement device (component 320 in Figures 13 and 14) and an alternate axial force generator. Furthermore, in the embodiment depicted in FIG. 11 , the conical spring 133 is moved to the output side of the transmission 300 and will switch to high.

Referring now to Figures 11 and 12, an alternative retainer 389 is disclosed. Holder 389 includes input and output stator plates 381a, 381b, however output stator plate 381b is omitted for ease of viewing. The output stator 381b of many embodiments is similar in structure to the input stator 381a. A plurality of stator curved plates 382 are attached to the stator discs 381a, 381b and have a first side facing the ball 1 and a second side facing away from the ball 1 . The second side 391 of each of the stator curved plates 382 is a flat surface that lies flat against one of the stator plates 381a, 381b. The stator curved plate 382 has two through holes for attaching it to the stator discs 381a, 381b with conventional fasteners or other type of attachment means. The stator curved plates 382 have rectangular slots on each of their first sides through which a plurality of flat spacers 383 are inserted to connect the stator 381 . The flat spacers 383 are used to set the distance between the stators 381, establish a solid connection between the stators 381, and ensure that the stators 381 are parallel and aligned.

The described design uses a substantially flat stator disc 181 . Thus, the stator disc 181 can be manufactured using a substantially flat sheet of rigid material. The stator disc 181 may be produced by any number of inexpensive manufacturing techniques such as stamping, fine blanking, or other techniques known in the industry. The stator disc 181 of this design can be manufactured from thin metal or sheet metal, plastic, ceramic, wood or paper products or any other material. The described design allows for a substantial reduction in material costs and allows the fabrication of relatively inexpensive components (with reasonably high tolerances).

Referring now to Figures 11, 13 and 14, an alternative breakaway device 320 is disclosed. FIG. 13 is a cross-sectional view taken near the axis of the transmission 300 and FIG. 14 is a cross-sectional view taken from the upper and outer sides of the transmission 300 generally radially inward toward the center. The ratchet 126 and ratchet bevel 127 of the previous embodiments are combined in this embodiment into one pawl gear 326 that engages the pawl 124 and has a bevel gear 328 interlocked. In other embodiments, bevel gear 328 may have non-bevel gear teeth. The clutch lever 322 is a rigid, flat L-shaped component with three or more holes. The centermost hole at the junction of the two legs forming the "L" shape positions the trip lever 322 rotatably and coaxially about the preloader 123 . A hole near the end of the long leg near the trip lever 322 allows insertion of the pawl pin 125 and allows the attachment to the pawl 124 to engage. A hole near the end of the end leg near the clutch lever 322 mates with the input disc connector 321 and receives and retains a clutch pin 329 that fits into a slot of the input disc connector 321 . Input disc connector 321 is rigidly attached to input disc 34 and has a slot for sliding engagement with clutch pin 329 . Otherwise, the operation of the alternative disengagement mechanism 320 is the same as that of the coasting mechanism 120 described in FIGS. 1 and 9 .

Referring now to FIGS. 11 and 15 , an alternative axial force generator includes a generally tapered wedge 360 located on the central axis of the transmission 300 and movable axially therealong. The tapered wedge 360 also cooperates with the spline 109 . When the transmission 300 is shifted low, the tapered wedge 360 is engaged by the follower rod 171 and moves axially in the same direction as the follower rod 171 . A tapered wedge 360 contacts a first end of an AFG (Axial Force Generator) control rod 362 near the axis of the transmission 300 . AFG lever 362 is an elongate member having a first semicircular end that engages tapered wedge 360 and extends radially outward from longitudinal axis 11 to a second end that engages input disc lever 112 . The AFG lever 362 is attached to the spline 109 with a pivot pin 361 about which the AFG lever 362 rotates. The pivot pin 361 is pivotally connected to the AFG lever 362 such that the second end of the AFG lever 362 engages the input disc lever 112 . The input disc control rod 112 is operatively attached to the bearing disc 60 and rotates thereabout such that the peripheral ramp 61 engages, thus switching the input torque from the screw 35 to the peripheral ramp 61 . Otherwise, the operation of the alternative axial force generator 360 is the same as the previous axial force generators of FIGS. 4 and 5 , seen and described therein.

Referring now to Figures 16 and 17, an alternate embodiment of the transmission 100 of Figure 1 is disclosed. For simplicity, only the differences between transmission 1700 in FIG. 17 and transmission 100 in FIG. 1 will be described. The transmission 100 in FIG. 1 includes a variator. The term variator may be used to describe a component of transmission 100 that changes the ratio of input to output speed. The components and assemblies comprising the transmission 401 of this embodiment include ball/leg member 403 in FIG. 5 , input disc 34 , output disc 101 , idler member 402 in FIG. 4 and retainer 89 in FIG. 7 . It should be noted that all components and components of transmission 401 may be varied to best suit a particular application of transmission 1700 , and that the general form comprising the above components and components of transmission 401 is depicted in FIG. 16 .

Transmission 1700 depicted in FIG. 17 is similar in assembly to transmission 100 , but includes two variators 401 . This configuration is beneficial in applications requiring high torque capacity in a transmission 1700 having a small diameter or overall size. This configuration also eliminates the need for radial bearings to support the bearing disc 114 and the output disc 101, thereby increasing overall efficiency. Since transmission 1700 has two transmissions 401, each transmission 401 has an output side, and transmission 1700 also has an output side. Thus, with three output sides, no conventional or similarly labeled components with "a" and "b" are used in this structure to distinguish input and output sides. However, as shown in FIG. 17, the right side is the input side and the left side is the output side.

Referring to Figures 17 to 19, the housing 423 surrounding and enclosing the transmission 1700 is depicted. Housing 423 is generally cylindrical and protects transmission 1700 from contact with outside components and from getting dirty, and additionally contains lubricant for proper operation. Shell 423 is attached to an engine, frame or other rigid body (not shown) with standard fasteners (not shown) installed through shell holes 424 . The housing 423 is open on its input side (the side with the housing aperture 424 or the right side as shown) to receive input torque. Input torque is transmitted from an external source to an input drive shaft 425 which may be a long, rigid rod or shaft capable of transmitting torque. The input drive shaft 425 transmits torque to the bearing plate 428 via splines, keying or other means. Bearing disc 428 is a disc-shaped rigid assembly capable of absorbing substantial axial forces generated by transmission 1700 and is similar in design to bearing disc 60 in FIG. 1 . Between flange 429 on the input end of input shaft 425 and bearing disc 428 , input shaft bearing 426 is coaxially positioned above input shaft 425 to allow a small amount of movement between bearing disc 428 and input shaft 425 . As the bearing disc 429 begins to rotate, the peripheral ramp 61 , ramp bearing 62 , bearing retainer 63 , input disc ramp 64 and output disc 34 rotate as previously described. This causes the ball 1 in the first transmission (the transmission on the input side) to spin.

At the same time, the second input disc 431 rotates with the rotation of the input shaft 425 . The second input disc 431 is rigidly attached to the input shaft 425 and can be keyed with a stop nut pressed onto the input shaft 425 (or by welding, pins or other attachment method). The second input disk 431 is located on the output side of the transmission device 1700 and is opposite to the bearing disk 428 . The second input disc 431 and bearing disc 428 absorb a substantial amount of the axial force generated by the peripheral ramp 61, ramp bearing 62 and input disc ramp 64 as a normal force to prevent slippage at the ball/disk interface (as previously described) . The second input disc 431 is similar in shape to the aforementioned input disc 34 and relies on the rotation of the input shaft 425 ; the second input disc 431 rotates the balls 1 in the second transmission 422 . The second transmission 422 is generally the opposite of the first transmission 420 and is further from the input side of the transmission 1700 such that the first transmission 420 is located between the second transmission 422 and the input side.

As mentioned above, the balls 1 in the first transmission 420 cause the output disc 430 to rotate through their rolling contact with the above-mentioned components. Although serving the same function as the output disc 101 as previously described, the output disc 430 has two opposing contact surfaces and is in contact with the balls 1 in the two derailleurs 420 , 422 . From the sectional view shown in FIG. 17 , the shape of the output disc 430 can be a shallow arch or an inverted shallow "V" shape, and its ends have contact surfaces with the balls 1 of the two transmissions. The output disc 430 is wound around the second transmission 422 and extends toward the output side generally in a cylindrical shape. In the depicted embodiment, the cylindrical shape of the output disc 430 continues towards the output side of the transmission 1700 and encircles the second input disc 431, after which the output disc 430 reduces in diameter before exiting the housing. At 423, it becomes an ordinary cylindrical shape with a small diameter again. To keep the output disk 430 concentric and aligned with the first and second input disks 34, 431, annular bearings 434, 435 may be used to radially align the output disk 431. A housing bearing 434 is located in the bore of the housing 423 above the output disc 430 and an output disc bearing 435 is located in the bore of the output disc 430 above the input shaft 425 to provide additional support. The output disc 430 may be constructed from two segments that are joined to form the output disc 430 . This allows the second transmission 422 to be assembled within the cylindrical housing of the output disc 430 .

This can be accomplished by using two annular flanges along the major diameter of the output disc 430 as shown in FIG. 17 . In some embodiments, the annular flange is located generally midway along the major diameter of the output disc 430 . Referring now to Figures 17, 20 and 21, the ball shaft 433 of the transmission 1700 is similar to the ball shaft 3 previously described and performs the same function. Furthermore, the function of the ball shaft 433 is to tilt the ball 1 to change the speed ratio of the transmission 1700 . The ball shafts 433 are elongated on their respective output sides and extend through the wall of the output stator 435 . The output stator 435 is similar to the previously described output stator 80b, but a plurality of radial slots 436 penetrate the wall of the output stator 435 throughout. The slots 436 of the output stator 435 continue all the way through the wall of the output stator 435 such that a series of equally spaced radial slots 436 extend radially from near the bore at the center of the output stator 435 to the outer periphery. Ball axles 433 have iris rollers 407 located coaxially above their elongated output ends. Diaphragm roller 407 is generally a cylindrical wheel rotatable on ball axle 433 and is designed to fit within groove 411 of iris plate 409 . Diaphragm plate 409 is an annular disc or plate having a hole through its center and is mounted coaxially about longitudinal axis 11 of transmission 1700 . Diaphragm plate 409 is thicker than twice the thickness of each diaphragm roll 407 and has iris grooves 411 extending radially outward from the vicinity of the aforementioned holes to near the periphery of diaphragm plate 409 . As the diaphragm grooves 411 extend radially, their annular position also changes so that as the diaphragm plate 409 rotates annularly about the longitudinal axis 11, the diaphragm grooves 411 provide the function of a cam system along their respective lengths. In other words, the slots 411 spiral outward from near the hole in the center of the diaphragm plate 409 to respective locations near its perimeter.

Diaphragm rollers 407 are radial along their outer diameter and have rounded corners at their outer corners so that their diameter remains constant within groove 411 of diaphragm plate 409 as ball axis 433 tilts. Diaphragm plate 409 has a thickness sufficient to allow diaphragm roller 407 to remain within slot 411 of diaphragm plate 433 from both transmissions 402, 422 (at all shift ratios). Diaphragm slot 411 operates in the manner of a conventional diaphragm plate and allows ball shaft 433 to move radially inward or outward as diaphragm plate 409 rotates. Diaphragm plate 409 has a first side facing the first variator and a second side facing the second variator, and is coaxially circumscribed about the longitudinal axis 11 of the transmission 1700 and is located adjacent projections in the two output stators 435 part (the tube-mounted extension from where the output stator 435 extends). Through axial holes (not shown) in the bosses of the output stators 435, the two output stators 435 can be attached together with conventional fasteners. The bosses of the output stator 435 have a plurality of holes through their center and a plurality of holes disposed radially outward from the center. In some embodiments, the boss on the output stator 435 creates a space slightly wider than the diaphragm plate 409 to allow the diaphragm plate 433 to rotate freely, some embodiments use bearings between the boss and the diaphragm plate 409 to accurately The position of the diaphragm plate 409 between the output stators 435 is precisely controlled. An iris cable 406 is attached to a first side of the diaphragm plate 409 near its outer diameter and extends longitudinally from the connection point. The diaphragm cable 406 is routed through the output stator 435 of the first transmission 420 in a direction such that the diaphragm plate 409 is rotated when it is pulled. The diaphragm cable 406 passes through the housing 423 to the outside of the transmission 1700 after passing through an aperture near the outer periphery of the output stator 435, where the diaphragm cable 406 can control the transmission ratio. Near the outer diameter of the diaphragm plate 409 is attached an iris spring 408 . The diaphragm spring 408 is also attached to the output stator 435 of the second transmission 422 . Diaphragm spring 408 exerts a resilient force against rotation of diaphragm plate 409 by tension applied by diaphragm cable 406 . When the tension from the diaphragm cable 406 is released, the diaphragm spring 408 returns the diaphragm plate 409 to its equilibrium position. Depending on the application of the transmission 1700, the diaphragm plate 409 may be configured such that when the diaphragm cable 406 is pulled, the diaphragm plate 409 switches the transmission 1700 to a higher gear ratio, and when the tension on the diaphragm cable 406 is released, Diaphragm spring 408 switches transmission 1700 to a low ratio. As an alternative, the diaphragm plate 409 may be configured such that when the diaphragm cable 406 is pulled, the diaphragm plate 409 switches the transmission 1700 to a lower ratio, and when the tension on the diaphragm cable 406 is released, the diaphragm spring 408 Shift transmission 1700 to a high ratio.

Referring to FIGS. 16 and 17 , the embodiment of transmission 1700 having two variators 420 , 422 requires a high degree of accuracy in aligning the additional rolling elements of transmission 1700 . All rolling elements must be aligned with each other, otherwise, efficiency will be compromised and the useful life of the transmission 1700 will be reduced. During assembly, the input disc 34, the output disc 430, the second input disc 431 and the idler assembly 402 are aligned on the same longitudinal axis. Additionally, the retainer 410 (in these embodiments comprising the two retainers 89 joined by the output stator 435 as described above) must also be aligned on this longitudinal axis to accurately position the ball/leg assembly 403 . All rolling elements are positioned relative to the input shaft 425 in order to achieve the alignment described above simply and accurately. A first input stator bearing 440 and a second input stator bearing 444 are disposed in bores of the input stators 440 , 444 over the input shaft 425 to aid in aligning the retainer 410 . The output stator bearing 442 located in the bore of the output stator 435 and above the input shaft 425 is also aligned with the retainer 410 . The first guide bearing 441 is located in the hole of the first switching guide 13b and above the input shaft 425, and the second guide bearing 443 is located in the hole of the second switching guide 13b and on the input shaft 425, so that the first and second The two idler assemblies 402 are aligned.

Referring to FIGS. 18 and 19 , the retainer 410 is attached to the shell 423 with the aforementioned shell connector 383 mounted to the shell slot 421 . Shell slot 421 is a longitudinal slot in shell 423 that extends to its input side (the open side of shell 423 ). In this described embodiment, the shell is normally closed on the input side, which is not shown in FIG. 19 , but is open on the input side and has an otherwise cylindrical body extending radially from the shell 423. The mounting flange has a shell hole 424 for mounting the shell 423 on the flange. During assembly, transmission 1700 may be inserted into housing 423 wherein housing connector 383 aligns in housing slot 421 to resist torque applied to holder 410 and prevent holder 410 from rotating. Shell connector holes 412 in the shell 423 allow fasteners to be inserted into corresponding holes in the shell connector 383 to secure the retainer 410 to the shell 423 .

FIG. 22 depicts an alternative embodiment of the retainer 470 of the transmission 1700 . To reduce manufacturing costs, it is sometimes preferable to minimize the number of different parts being manufactured and to design parts that can be produced cheaply using mass production processes. The illustrated holder 470 uses four different low cost designed components with common fixtures to assemble these different components. The stator 472 is generally a flat disc shaped member and has a plurality of radial slots extending radially outward from near a central bore through which the input shaft 425 rotates. Ball shafts (item 433 in FIG. 17 ) extend through slots in the stator 472 . A plurality of holes 471 surrounding the central hole of the stator 472 are used to secure the stator 472 to other components. There are four stators 472 which in this embodiment are similar to each other and form part of the holder 470 . Two input stators 472 are located at each end of holder 470 and two output stators 472 are located near the center of holder 472 , the stators being rigidly attached together using stator bridges 477 .

Still referring to the embodiment depicted in FIG. 22 , the stator bridge 477 is a disc-shaped member having a central hole and having a through hole between the inner and outer diameters of the stator bridge 477 . The holes in the stator bridge 477 are complementary to the holes in the stator 472 to allow the stator 472 to be fixed to the stator bridge 477 . Diaphragm plates 409 (not shown) are located radially outboard of stator bridges 477 and axially between output stators 472 . In some embodiments, the stator bridge 477 is slightly thicker than the diaphragm plate 409 to allow the diaphragm plate 409 to rotate freely, while in other embodiments, between the output stator 472 and the diaphragm plate 409, and between the stator bridge 477 and the diaphragm There are bearings between the plates 409 . Thus, the outer diameter of the stator bridge 477 is used to locate the inner diameter of the diaphragm plate 409 and set the axis of the diaphragm plate 409 .

A spacer 473 joins the input stator 472 to the output stator 472 . In one embodiment, the spacers 473 are made of flat material (eg, sheet metal or plate) and then shaped to create their unique shape, which will serve several purposes. The spacers 473 are generally flat rectangular pieces having a hole 475 formed in their center and vertical extensions at each end. The spacer 473 sets the correct distance between the stators 472, forms the structural frame of the retainer 470 to prevent the balls 1 from orbiting along the longitudinal axis of the transmission 1700, aligns the stator holes with each other so that the centers of the stators 472 are aligned and the stator The angular orientation of 472 is the same, prevents the retainer 470 from twisting or tipping, and provides a rolling concave surface 479 on which the stator wheel 30 rolls. Each spacer 473 is formed with two ends that are bent out of the plane of the rest of the spacer to form the mounting area 480 and the curved surface 479 of the holder 470 . The spacer 473 has mounting holes 481 in the side in contact with the stator 472 that line up with corresponding holes in the stator 472 to allow the spacer 473 to be fixed to the stator 472 . A hole 475 near the center of the spacer 473 provides clearance for the ball 1 .

In one embodiment, each ball 1 has two spacers 473, although more or fewer spacers 473 could be used. Each spacer 473 is paired back-to-back with the spacer in its counterpart to form an I-beam shape. In one embodiment, rivets 476 may be used to connect spacer 473 to stator 472 and stator 472 to stator bridge 477 . During assembly, rivets 476 are threaded securely into the bores of stator 472 , spacer 473 and stator bridge 477 . Only two rivets 476 are depicted in Figure 22, but all rivets could use the same design. The spacer 473 used in the first transmission 420 also has a housing connector 474 that extends generally radially outward from the spacer 473 and then bends generally perpendicularly. Shell connector 474 in some embodiments is made from a flat material, such as sheet metal, which is die stamped and then formed into a final shape. Shell connector 474 may be integrally formed with or rigidly attached to spacer 473 and extend radially into shell 423 between input disk 34 and output disk 430 . In some embodiments, shell connector 474 is formed as part of spacer 473 during manufacture of spacer 473 . Shell connector holes 478 in the vertical end of the shell connector 474 line up with corresponding shell connector holes (item 412 in FIG. 19 ), thereby enabling the retainer 470 to be secured to the shell 423 with standard fasteners.

The design shown in Figure 22 uses a stator plate 472 that is substantially flat and can be fabricated using a substantially flat sheet of rigid material. Additionally, the spacer 473 with and without the shell connector 474 is also substantially flat and can be formed from a flat sheet of material, although in many embodiments the vertical ends of the shell connector 474, the mounting area 480, and the curved surface 480 Formed in successive bending steps. The stator disks 472 and spacers may be formed by a variety of inexpensive manufacturing processes, such as stamping, fine blanking, or other processes known in the art. The stator plate 472 and spacer 473 of this design can be made of thin metal sheet, plastic, ceramic, wood or paper or any other material. As described above with reference to Figure 12, the design shown reduces material costs and reduces the manufacturing costs of manufacturing these inexpensive components to suitably high tolerances. Furthermore, while the embodiment shown in FIG. 22 shows a dual chamber design for the transmission, the components fabricated by the inexpensive process described above could also be used in a single chamber design for the retainer 470 . As one example, two of the stator plates 472 shown may be attached to a spacer 473 (with shell connection 474 on the right of FIG. 22 ) to create a single chamber design for use with the embodiments described herein.

FIG. 23 depicts an embodiment of a ball 1 for use with the transmission 100 , 1700 of FIGS. 1 and 17 . The ball 1 has helical grooves 450 which suck lubricant through the ball 1 . In one embodiment, two helical grooves 450 are used, starting at the end of the hole in the ball 1 and continuing through the other end of the hole. The helical groove 450 carries lubricant through the ball 1 to remove heat and provides lubrication between the ball 1 and the ball shaft 3, 433 to improve efficiency and extend the life of the transmission 100, 1700.

FIG. 24 depicts an alternative leg 460 to the ball/leg assembly 430 of FIG. 5 . Compared with the leg 2 of FIG. 5 , the leg 460 is simplified and does not have the stator wheel 30 , the stator wheel pin 31 , the guide wheel 21 or the guide wheel pin 22 . The legs 460 have a raised surface on the side of the first leg facing away from the ball 1 which fits into a corresponding groove (not shown) of each stator 80 . On the second side 465 facing the ball 1, the leg 460 is concave and has a convex curve formed in the vicinity of its radially inner end with a leg cam 466 in contact with the switching guide 13 and Axially and radially set by switching guides 13 . The transverse and longitudinal lubricant channels 460, 464, respectively, allow lubricant to be injected into the legs and delivered to different areas. The lubricant is used to cool the legs and other parts of the transmission 100 , 1700 , and also to minimize friction where the legs come into contact with the switching guide 13 and the stator 80 . It should be noted that additional channels may be drilled or formed in the legs 460 to route lubricant to other areas, any channel openings may be used as lubricant inlets. Longitudinal channel 464 is a hole running the length of leg 460 , generally in the center, and extending through the bottom and through ball axle hole 461 at the top of each leg 460 . The transverse channel 462 is a blind bore formed generally perpendicular to the longitudinal channel 464 and extends beyond the first leg side 463 . In some embodiments as described, transverse channel 462 intersects and terminates longitudinal channel 464 without penetrating second leg side 465 . In embodiments where the transverse channel 462 intersects the longitudinal channel 464 , lubricant may enter at the opening of the transverse channel 462 and then pass through the channel 464 .

In some embodiments, the ball shaft 3 , 433 is press fit in the ball 1 and rotates with the ball 1 . The ball shaft 3 , 433 rotates inside the ball shaft hole 461 and in the roller 4 . The lubricant flows through the top of the leg 460 into the ball axle bore 1 , where it provides a liquid layer to reduce friction.

Referring to Figures 25-27, a graphical method for estimating the convex curve 97 on the switching guide 13 is disclosed. For purposes of simplicity, the idler 18 , idler bearing 17 and shift guide 13 are combined to simplify the analysis and description of the correct curve plate 97 of the shift guide 13 of one embodiment. For description and analysis, the following assumptions are made:

1. The center of the ball 1 is fixed so that the ball 1 can rotate around its axis, and its axis can rotate, but the ball 1 has no displacement.

2. Ball 1, ball shaft 3, 343, leg 2 and guide wheel 21 rotate as a rigid body.

3. The idler 18 is only able to move in the x direction.

4. The outer peripheral surface of the idler 18 is tangent to the circumference of the ball 1 .

5. The side of the switching guide 13 is tangent to the circumference of the guide wheel 21 .

6. Angular rotation of the ball 1 causes a linear movement of the switching guide 13 and vice versa.

7. When the ball axis 3, 343 is horizontal or parallel to the longitudinal axis 11, the contact point of each guide wheel 21 and its respective switching guide 13 is at the starting point of the curved plate 97, where the switching guide 13 The vertical walls transition into curved panels 97 . When the ball 1 is tilted, only one guide wheel 21 is in contact with the curved plate 97; the other guide wheels 21 are in contact with the vertical wall of the switching guide 13.

The goal of this analysis is to find the approximate coordinates (as a function of the angle of inclination of the axis of the ball 1 ) of the point where the guide wheel 21 comes into contact with the curved plate 97 on the switching guide 13 . If the coordinates of the different ball axes 3 , 343 are plotted, a curve can be fitted by the coordinate points following the path of the guide wheel 21 /switching guide 13 contact point for all switching ranges.

The above-mentioned coordinates start at the initial position (x0, y0) of guide wheel 21/switching guide 13 contact when the angle of rotation is zero, and then change at each increasing angular position during the inclination of the ball 1 . By comparing these coordinates, the position (xn, yn) of the guide wheel 21 /switching guide 13 contact is determined as a function of the angle of inclination (theta) of the ball 1 .

From Figure 25 and Figure 26, the known variables are:

1. H1: The vertical distance from the center of the ball 1 to the center of the guide wheel 21 .

2. H2: The sum of the radius of the ball 1 and the radius of the idler 18.

3. W: the horizontal distance from the center of the ball 1 to the center of the guide wheel 21 .

4.rw: guide wheel radius.

From these known variables, the following relationship can be determined:

$\mathrm{<span\; class="notranslate">\; <figure-callout\; id="1"\; label="R"\; filenames="C200480011347E01051.png,C200480011347E01061.png"\; state="\{\{state\}\}">R</figure-callout></span>}\mathrm{<span\; class="notranslate">\; 1</span>}<span\; class="notranslate">\; =</span>{<span\; class="notranslate">\; [</span>{<span\; class="notranslate">\; (</span>\mathrm{<span\; class="notranslate">\; W</span>}<span\; class="notranslate">\; -</span>\mathrm{<span\; class="notranslate">\; rw</span>}<span\; class="notranslate">\; )</span>}^{\mathrm{<span\; class="notranslate">\; 2</span>}}<span\; class="notranslate">\; +</span>\mathrm{<span\; class="notranslate">\; <figure-callout\; id="1"\; label="h"\; filenames="C200480011347E01051.png,C200480011347E01061.png"\; state="\{\{state\}\}">h</figure-callout></span>}{\mathrm{<span\; class="notranslate">\; 1</span>}}^{\mathrm{<span\; class="notranslate">\; 2</span>}}<span\; class="notranslate">\; ]</span>}^{<span\; class="notranslate">\; ^</span>}<span\; class="notranslate">\; (</span>\frac{\mathrm{<span\; class="notranslate">\; 1</span>}}{\mathrm{<span\; class="notranslate">\; 2</span>}}<span\; class="notranslate">\; )</span><span\; class="notranslate">\; -</span><span\; class="notranslate">\; -</span><span\; class="notranslate">\; -</span><span\; class="notranslate">\; (</span>\mathrm{<span\; class="notranslate">\; 1</span>}<span\; class="notranslate">\; )</span>$

Phi＝TAN ^-1 [(W-rw)/H1] (2)

x0＝W-rw (3)

y0＝H1-H2 (4)

BETA＝TAN ^-1 (H1/W) (5)

$\mathrm{<span\; class="notranslate">\; <figure-callout\; id="2"\; label="R"\; filenames="C200480011347E01021.png,C200480011347E01051.png"\; state="\{\{state\}\}">R</figure-callout></span>}\mathrm{<span\; class="notranslate">\; 2</span>}<span\; class="notranslate">\; =</span>{<span\; class="notranslate">\; [</span>\mathrm{<span\; class="notranslate">\; <figure-callout\; id="1"\; label="h"\; filenames="C200480011347E01051.png,C200480011347E01061.png"\; state="\{\{state\}\}">h</figure-callout></span>}{\mathrm{<span\; class="notranslate">\; 1</span>}}^{\mathrm{<span\; class="notranslate">\; 2</span>}}<span\; class="notranslate">\; +</span>{\mathrm{<span\; class="notranslate">\; W</span>}}^{\mathrm{<span\; class="notranslate">\; 2</span>}}<span\; class="notranslate">\; ]</span>}^{<span\; class="notranslate">\; ^</span>}<span\; class="notranslate">\; (</span>\frac{\mathrm{<span\; class="notranslate">\; 1</span>}}{\mathrm{<span\; class="notranslate">\; 2</span>}}<span\; class="notranslate">\; )</span><span\; class="notranslate">\; -</span><span\; class="notranslate">\; -</span><span\; class="notranslate">\; -</span><span\; class="notranslate">\; (</span>\mathrm{<span\; class="notranslate">\; 6</span>}<span\; class="notranslate">\; )</span>$

At this time, it is assumed that the ball 1 is inclined at an angle THETA, which causes the switching guide 13 to move in the x-axis direction (see FIG. 26 ). Therefore, the following equation can be derived:

Nu＝90°-BETA-THETA (7)

x2＝R2*SIN(Nu) (8)

x3＝x2-rw (9)

x_switch boot device = x0-x3 (10)

This is the x distance that the switching guide 13 moves for a given THETA.

x4＝R1*SIN(Phi+THETA) (11)

x_guide wheel=x4-x0 (12)

This is the x distance that the guide wheel 21 moves for a given THETA.

At this point, it is convenient to determine the x'-y' origin at the center of the idler 18. This is useful for plotting guide wheel 21/switching guide 13 contact coordinates.

x1=x0-(x_switch guide device-x_guide wheel) (13)

By comparing equations (10), (12) and (13),

x1＝x4+x3-x0 (14)

This is the x' position of the guide wheel 21/switching guide 13 contact.

Finding the y' position where the guide wheel 21/switch guide 13 contacts is relatively simple,

y2＝R1*COS(Phi+THETA) (15)

y1=H2-y2 (16)

This is the y' position of the guide wheel 21/switching guide 13 contact.

Thus, x1 and y1 can be determined and then described for different THETA values. This is shown in Figure 27. Since the coordinates have already been determined, it is simple for most CAD programs to fit the curve through the above coordinates. Methods of curve fitting may include any suitable algorithm, such as linear regression, to determine the appropriate curve for the above relationship; although direct functions derived from the relationships described above can also be developed.

Referring now to FIGS. 1 , 7 and 28 , the transmission 100 may be used as a continuously variable planetary gear set 500 . Referring to Figures 1 and 7, in an embodiment where the retainer 89 is free to rotate about the longitudinal axis 11, the idler gear 18 acts as a sun gear, the balls 1 act as planetary gears, and the retainer 89 retains the balls 1 and acts as a planetary gear. The role of the gear support, the input disc 34 is the first ring gear, and the output disc 101 is the second ring gear. Each ball 1 is in contact with the input disc 34 , the output disc 101 and the idler 18 and is supported or held in radial position by the retainer 89 .

FIG. 28 is a skeleton diagram or schematic diagram of the planetary gear set 500 , wherein only the upper half of the planetary gear set 500 is shown for simplicity. The figure is cut at the centerline of the planetary gear set 500 or the longitudinal axis 11 of the transmission 100 . The line of contact formed by the output disk 101 around each ball 1 forms a variable rolling diameter, which allows each ball 1 part to operate as a first planetary gear 501 . The contact between the balls 1 and the idler 18 establishes a variable rolling diameter, which allows each part of the ball 1 to operate as a second planetary gear 502 . The contact between the balls 1 and the input disc 34 establishes a variable rolling diameter, which allows each part of the balls 1 to function as a third planetary gear 503 .

In the planetary gear set 500 embodiment, those skilled in the art will recognize that different radius and thrust bearings may be beneficially used to maintain the position of the input disc 34, output disc 101 and retainer 89 relative to each other. Those skilled in the art will also recognize that solid or hollow shafts may be used and attached to the input disc 34, output disc 101, retainer 89, and/or idler 18 to properly perform the functions described herein, such modifications It will be apparent to those skilled in the art of rotary power transmission.

Referring now to FIGS. 29-31 , the respective diameters of the first planetary gear 501 , the second planetary gear 502 and the third planetary gear 503 may be varied by the transmission 100 . Figure 29 shows a transmission 100 having first and third planet gears 501, 503 of the same diameter with the second planet gear 502 at its largest diameter. By tilting the ball 1 as previously described, the diameter of the planet gears 501 , 502, 503 is changed, thereby changing the input to output speed of the transmission 1700 . Figure 30 shows the balls tilted so that the diameter of the first planetary gear 501 increases and the diameter of the second planetary gear 502 and third planetary gear 503 decreases. Fig. 31 shows balls that are tilted such that the diameter of the third planetary gear 503 increases and the diameters of the first planetary gear 501 and the second planetary gear 502 decrease.

By varying the torque source between input disc 34 , idler 18 and/or retainer 89 many different speed combinations are possible. Additionally, some implementations use more than one input. For example, both input disc 34 and retainer 89 can provide input torque and can rotate at the same or different speeds. One or more input torque sources can be variable speed to increase the ratio potential of the transmission 100 . The list given below lists some of the combinations obtained by using the transmission 100 as a planetary gear set. In this listing, a source of input torque (or "input") is indicated by an "I", an output is indicated by an "O", and a component that is fixed so as not to rotate about the longitudinal axis 11 is indicated by an "F". If the component Free spins are allowed, indicated by an "R". "Single-in/single-out" is used to mean having one input and one output. "Dual input/single output" is used to mean having two inputs and one output. "Single input/dual output" is used to mean having one input and two outputs. "Dual-in/dual-out" is used to mean having two inputs and two outputs. "Triple input/single output" is used to mean having three inputs and one output. "Single input/three output" is used to mean having one input and three outputs.

Configuration Input Disc (34) Idler (18) Retainer (89) Output Disc (101) Single input/single output F I F O Single input/single output R I F O Single input/single output F I R O Single input/single output R I R O Single input/single output F I O F Single input/single output R I O F Single input/single output F I O R Single input/single output R I O R Single input/single output I R F O Single input/single output I F R O

Single input/single output I F F O Single input/single output I R R O Single input/single output I F O F Single input/single output I F O R Single input/single output I R O F Single input/single output I R O R Single input/single output F F I O Single input/single output F R I O Single input/single output R F I O Single input/single output R R I O Single input/single output F O I F Single input/single output R O I F Single input/single output F O I R Single input/single output R O I R

configuration Input Disk(34) Idler (18) Holder(89) Output tray(101) Dual input/single output I I f o Dual input/single output I I R o Dual input/single output I I o f Dual input/single output I I o R Dual input/single output I o I f Dual input/single output I o I R Dual input/single output I f I o Dual input/single output I R I o Dual input/single output f I I o Dual input/single output R I I o

configuration Input Disk(34) Idler (18) Holder(89) Output tray(101) single input/dual output I o f o single input/dual output I o R o

single input/dual output I f o o single input/dual output I R o o single input/dual output I o o f single input/dual output I o o R single input/dual output f I o o single input/dual output R I o o single input/dual output f o I o single input/dual output R o I o

configuration Input Disk(34) Idler (18) Holder(89) Output tray(101) Dual Input/Dual Output I I o o Dual Input/Dual Output I o I o

configuration Input Disk(34) Idler (18) Holder(89) Output tray(101) Triple input/single output I I I o Triple input/single output I I o I

configuration Input Disk(34) Idler (18) Holder(89) Output tray(101) single input/three output I o o o

Referring now to FIG. 32 , the transmission 100 can also be combined by parallel power paths with planetary gear sets 505 to create more speed combinations. A typical planetary gear set 505 includes a centrally located sun gear, distributed around and engaged with the sun gear, a plurality of planet gears (all rotatably attached at their respective centers to planet gear carriers (often referred to simply as carrier)) and the ring gear that surrounds and engages the planetary gears. Many speed combinations are available by switching the input torque and output source between sun gear, carrier and ring gear. The planetary gearset 505 in combination with the transmission 100 provides a very large number of speed combinations and in some cases an infinitely variable transmission can be obtained. In FIG. 32, the torque input of transmission 100 is coupled to input disc 34 and first gear 506, which is generally coaxial with input disc 34 and contacts and rotates second gear 509 to drive a parallel power path. The basic configuration that couples the input disc 34 of transmission 100 or CVT 100 and the input of the parallel power path to a prime mover or other source of torque (eg, an electric motor or other power-generating device) is referred to as "input coupling." By changing the diameter of the first gear 506 and the second gear 509, the input speed to the parallel power paths can be changed. Second gear 509 is attached to and rotates gear shaft 508 , which in some embodiments rotates gearbox 507 . Gearbox 507 (implemented as a design option in these embodiments) can further vary the rotational speed of the parallel power paths, and can be conventional gearing. The gearbox 507 rotates the gearbox shaft 511 which in turn rotates the third gear 510 . In embodiments where gearbox 507 is not used, gear shaft 508 drives third gear 510 . The third gear 510 drives the sun gear, carrier or ring gear of the planetary gear set 505 and has a diameter capable of producing the desired speed/torque ratio. As an alternative, the third gear 510 can be eliminated and the gearbox shaft 508 can directly drive the sun gear, carrier or ring gear of the planetary gear set 505 . The planetary gear set 505 also has an input from the output of the CVT 100, which drives the other sun gear, carrier or ring gear.

In the table below, many, if not all, of the various input and output combinations titled "Input Coupling" (possibly having the basic arrangement described immediately above) are identified. In this table, "IT" denotes the input torque source into the CVT 100, "I1" denotes the planetary gear set 505 coupled to the output of the CVT 100, and "OV" denotes the planetary gear set 505 coupled to the output of the vehicle or machine "F" denotes a component of planetary gear set 505 or transmission 100 that is fixed so as not to rotate about its axis, "I2" denotes a component coupled to a parallel path, which is third gear 509, and "R" denotes a component that can A component that rotates freely about its axis so that it does not drive other components. For this table and the table below it (titled "Output Coupling"), it is assumed that the ring gear is the only planetary gear set 505 that is fixed, in order to reduce the total number of tables that must be presented in this text. The sun gear or planet gear carrier can also be fixed with corresponding input and output combinations for other components, which are not given here for the purpose of reducing the description, but can be easily determined based on the two tables below.

transmission Input Disk(34) Idler (18) Holder(89) Output tray(101) ring gear bracket sun gear single input/single output f IT f o I1 I2 OV f IT f o OV I1 I2 f IT f o I2 OV I1 f IT f o f I1 I2, OV f IT f o f I2, OV I1 f IT f o f I2 I1, OV f IT f o f I1, OV I2 single input/single output R IT f o I1 I2 OV R IT f o OV I1 I2 R IT f o I2 OV I1 R IT f o f I1 I2, OV R IT f o f I2, OV I1 R IT f o f I2 I1, OV R IT f o f I1, OV I2 single input/single output f IT R o I1 I2 OV f IT R o OV I1 I2 f IT R o I2 OV I1 f IT R o f I1 I2, OV f IT R o f I2, OV I1 f IT R o f I2 I1, OV f IT R o f I1, OV I2 single input/ R IT R o I1 I2 OV

single output R IT R o OV I1 I2 R IT R o I2 OV I1 R IT R o f I1 I2, OV R IT R o f I2, OV I1 R IT R o f I2 I1, OV R IT R o f I1, OV I2 single input/single output f IT o f I1 I2 OV f IT o f OV I1 I2 f IT o f I2 OV I1 f IT o f f I1 I2, OV f IT o f f I2, OV I1 f IT o f f I2 I1, OV f IT o f f I1, OV I2 single input/single output R IT o f I1 I2 OV R IT o f OV I1 I2 R IT o f I2 OV I1 R IT o f f I1 I2, OV R IT o f f I2, OV I1 R IT o f f I2 I1, OV R IT o f f I1, OV I2 single input/single output f IT o R I1 I2 OV f IT o R OV I1 I2 f IT o R I2 OV I1 f IT o R f I1 I2, OV f IT o R f I2, OV I1 f IT o R f I2 I1, OV

f IT o R f I1, OV I2 single input/single output R IT o R I1 I2 OV R IT o R OV I1 I2 R IT o R I2 OV I1 R IT o R f I1 I2, OV R IT o R f I2, OV I1 R IT o R f I2 I1, OV R IT o R f I1, OV I2 single input/single output IT R f o I1 I2 OV IT R f o OV I1 I2 IT R f o I2 OV I1 IT R f o f I1 I2, OV IT R f o f I2, OV I1 IT R f o f I2 I1, OV IT R f o f I1, OV I2 single input/single output IT f R o I1 I2 OV IT f R o OV I1 I2 IT f R o I2 OV I1 IT f R o f I1 I2, OV IT f R o f I2, OV I1 IT f R o f I2 I1, OV IT f R o f I1, OV I2 single input/single output IT f f o I1 I2 OV IT f f o OV I1 I2 IT f f o I2 OV I1 IT f f o f I1 I2, OV

IT f f o f I2, OV I1 IT f f o f I2 I1, OV IT f f o f I1, OV I2 single input/single output IT R R o I1 I2 OV IT R R o OV I1 I2 IT R R o I2 OV I1 IT R R o f I1 I2, OV IT R R o f I2, OV I1 IT R R o f I2 I1, OV IT R R o f I1, OV I2 single input/single output IT f o f I1 I2 OV IT f o f OV I1 I2 IT f o f I2 OV I1 IT f o f f I1 I2, OV IT f o f f I2, OV I1 IT f o f f I2 I1, OV IT f o f f I1, OV I2 single input/single output IT f o R I1 I2 OV IT f o R OV I1 I2 IT f o R I2 OV I1 IT f o R f I1 I2, OV IT f o R f I2, OV I1 IT f o R f I2 I1, OV IT f o R f I1, OV I2 single input/single output IT R o f I1 I2 OV IT R o f OV I1 I2

IT R o f I2 OV I1 IT R o f f I1 I2, OV IT R o f f I2, OV I1 IT R o f f I2 I1, OV IT R o f f I1, OV I2 single input/single output IT R o R I1 I2 OV IT R o R OV I1 I2 IT R o R I2 OV I1 IT R o R f I1 I2, OV IT R o R f I2, OV I1 IT R o R f I2 I1, OV IT R o R f I1, OV I2 single input/single output f f IT o I1 I2 OV f f IT o OV I1 I2 f f IT o I2 OV I1 f f IT o f I1 I2, OV f f IT o f I2, OV I1 f f IT o f I2 I1, OV f f IT o f I1, OV I2 single input/single output f R IT o I1 I2 OV f R IT o OV I1 I2 f R IT o I2 OV I1 f R IT o f I1 I2, OV f R IT o f I2, OV I1 f R IT o f I2 I1, OV f R IT o f I1, OV I2 single input/ R f IT o I1 I2 OV

single output R f IT o OV I1 I2 R f IT o I2 OV I1 R f IT o f I1 I2, OV R f IT o f I2, OV I1 R f IT o f I2 I1, OV R f IT o f I1, OV I2 single input/single output R R IT o I1 I2 OV R R IT o OV I1 I2 R R IT o I2 OV I1 R R IT o f I1 I2, OV R R IT o f I2, OV I1 R R IT o f I2 I1, OV R R IT o f I1, OV I2 single input/single output f o IT f I1 I2 OV f o IT f OV I1 I2 f o IT f I2 OV I1 f o IT f f I1 I2, OV f o IT f f I2, OV I1 f o IT f f I2 I1, OV f o IT f f I1, OV I2 single input/single output R o IT f I1 I2 OV R o IT f OV I1 I2 R o IT f I2 OV I1 R o IT f f I1 I2, OV R o IT f f I2, OV I1 R o IT f f I2 I1, OV

R o IT f f I1, OV I2 single input/single output f o IT R I1 I2 OV f o IT R OV I1 I2 f o IT R I2 OV I1 f o IT R f I1 I2, OV f o IT R f I2, OV I1 f o IT R f I2 I1, OV f o IT R f I1, OV I2 single input/single output R o IT R I1 I2 OV R o IT R OV I1 I2 R o IT R I2 OV I1 R o IT R f I1 I2, OV R o IT R f I2, OV I1 R o IT R f I2 I1, OV R o IT R f I1, OV I2

Referring to the embodiment shown in FIG. 33, the torque input source drives the coupled planetary gear set 505 as input to the CVT 100. One or more components of CVT 100 are coupled to the parallel power path and output of the transmission. In this embodiment, the parallel power path is as follows: A component of planetary gear set 505 (sun gear, carrier or ring gear) meshes with third gear 510, which turns gear shaft 508, which in turn drives the aforementioned Gearbox 507. The gearbox 507 rotates the gearbox shaft 511 which rotates the second gear 509 which in turn drives the first gear 506 . The first gear 506 is mounted on the output shaft of the transmission, which is also coupled to the output of the CVT 100. In this embodiment, the planetary gear set 505 is coupled to the transmission's torque source and provides torque to the parallel path and the CVT 100 from which torque is coupled at the vehicle and equipment output. If the planetary gearset 505 is so coupled to provide torque to the CVT 100 and a fixed ratio parallel path, the two paths are coupled at the output (e.g., in a drive shaft, wheel, or other load device), this configuration is called " Output Coupling". In this basic configuration, the planetary gear set 505 in combination with the CVT 100 provides a very large number of speed combinations and in some cases an infinitely variable transmission can be obtained.

In the table below, many, if not all, possible combinations of the base unit are shown in FIG. 33 titled "Output Coupling". In the table, for planetary gear set 505, "O1" denotes the component of planetary gear set 505 that is coupled to CVT 100, "I" denotes the input from the engine, a human, or any other source, and "F" is fixed so that A component that cannot rotate about its own axis, and "02" refers to a component that is coupled to a parallel path via planetary gears 510. For the CVT 100, "I" refers to the component coupled to the planetary gear set 505, "O" refers to the component coupled to the output of the vehicle or machine, "F" refers to the fixed assembly as just described, and "R" refers to A component that rotates freely about its axis so that it does not drive any other components.

transmission ring gear bracket sun gear Input Disk(34) Idler (18) Holder(89) Output tray(101) single input/single output I O1 O2 f I f o O2 I O1 f I f o O1 O2 I f I f o f I, O1 O2 f I f o f O2 I, O1 f I f o f I, O2 O1 f I f o f O1 I, O2 f I f o single input/single output I O1 O2 R I f o

O2 I O1 R I f o O1 O2 I R I f o f I, O1 O2 R I f o f O2 I, O1 R I f o f I, O2 O1 R I f o f O1 I, O2 R I f o single input/single output I O1 O2 f I R o O2 I O1 f I R o O1 O2 I f I R o f I, O1 O2 f I R o f O2 I, O1 f I R o f I, O2 O1 f I R o f O1 I, O2 f I R o single input/single output I O1 O2 R I R o O2 I O1 R I R o O1 O2 I R I R o f I, O1 O2 R I R o f O2 I, O1 R I R o f I, O2 O1 R I R o f O1 I, O2 R I R o single input/single output I O1 O2 f I o f O2 I O1 f I o f O1 O2 I f I o f f I, O1 O2 f I o f f O2 I, O1 f I o f f I, O2 O1 f I o f f O1 I, O2 f I o f

single input/single output I O1 O2 R I o f O2 I O1 R I o f O1 O2 I R I o f f I, O1 O2 R I o f f O2 I, O1 R I o f f I, O2 O1 R I o f f O1 I, O2 R I o f single input/single output I O1 O2 f I o R O2 I O1 f I o R O1 O2 I f I o R f I, O1 O2 f I o R f O2 I, O1 f I o R f I, O2 O1 f I o R f O1 I, O2 f I o R single input/single output I O1 O2 R I o R O2 I O1 R I o R O1 O2 I R I o R f I, O1 O2 R I o R f O2 I, O1 R I o R f I, O2 O1 R I o R f O1 I, O2 R I o R single input/single output I O1 O2 I R f o O2 I O1 I R f o O1 O2 I I R f o f I, O1 O2 I R f o f O2 I, O1 I R f o

f I, O2 O1 I R f o f O1 I, O2 I R f o single input/single output I O1 O2 I f R o O2 I O1 I f R o O1 O2 I I f R o f I, O1 O2 I f R o f O2 I, O1 I f R o f I, O2 O1 I f R o f O1 I, O2 I f R o single input/single output I O1 O2 I f f o O2 I O1 I f f o O1 O2 I I f f o f I, O1 O2 I f f o f O2 I, O1 I f f o f I, O2 O1 I f f o f O1 I, O2 I f f o single input/single output I O1 O2 I R R o O2 I O1 I R R o O1 O2 I I R R o f I, O1 O2 I R R o f O2 I, O1 I R R o f I, O2 O1 I R R o f O1 I, O2 I R R o single input/single output I O1 O2 I f o f O2 I O1 I f o f O1 O2 I I f o f

f I, O1 O2 I f o f f O2 I, O1 I f o f f I, O2 O1 I f o f f O1 I, O2 I f o f single input/single output I O1 O2 I f o R O2 I O1 I f o R O1 O2 I I f o R f I, O1 O2 I f o R f O2 I, O1 I f o R f I, O2 O1 I f o R f O1 I, O2 I f o R single input/single output I O1 O2 I R o f O2 I O1 I R o f O1 O2 I I R o f f I, O1 O2 I R o f f O2 I, O1 I R o f f I, O2 O1 I R o f f O1 I, O2 I R o f single input/single output I O1 O2 I R o R O2 I O1 I R o R O1 O2 I I R o R f I, O1 O2 I R o R f O2 I, O1 I R o R f I, O2 O1 I R o R f O1 I, O2 I R o R single input/single output I O1 O2 f f I o

O2 I O1 f f I o O1 O2 I f f I o f I, O1 O2 f f I o f O2 I, O1 f f I o f I, O2 O1 f f I o f O1 I, O2 f f I o single input/single output I O1 O2 f R I o O2 I O1 f R I o O1 O2 I f R I o f I, O1 O2 f R I o f O2 I, O1 f R I o f I, O2 O1 f R I o f O1 I, O2 f R I o single input/single output I O1 O2 R f I o O2 I O1 R f I o O1 O2 I R f I o f I, O1 O2 R f I o f O2 I, O1 R f I o f I, O2 O1 R f I o f O1 I, O2 R f I o single input/single output I O1 O2 R R I o O2 I O1 R R I o O1 O2 I R R I o f I, O1 O2 R R I o f O2 I, O1 R R I o f I, O2 O1 R R I o f O1 I, O2 R R I o

single input/single output I O1 O2 R o I o O2 I O1 f o I f O1 O2 I f o I f f I, O1 O2 f o I f f O2 I, O1 f o I f f I, O2 O1 f o I f f O1 I, O2 f o I f single input/single output I O1 O2 R o I f O2 I O1 R o I f O1 O2 I R o I f f I, O1 O2 R o I f f O2 I, O1 R o I f f I, O2 O1 R o I f f O1 I, O2 R o I f single input/single output I O1 O2 f o I R O2 I O1 f o I R O1 O2 I f o I R f I, O1 O2 f o I R f O2 I, O1 f o I R f I, O2 O1 f o I R f O1 I, O2 f o I R single input/single output I O1 O2 R o I R O2 I O1 R o I R O1 O2 I R o I R f I, O1 O2 R o I R f O2 I, O1 R o I R

f I, O2 O1 R o I R f O1 I, O2 R o I R

Referring to the embodiment shown in FIG. 23 , the combinations in the base input coupling arrangement with two torque input sources to planetary gearset 505 are shown in the table below entitled "Input Coupling Dual Input Power Paths". References in this table refer to the same components as in the table above, except for the planetary gear set 505, "I1" refers to the output of the CVT 100, and "I2" is coupled to the parallel path (in this embodiment is a component of the planetary gear 510).

transmission Input Disk(34) Idler (18) Holder(89) Output tray(101) ring gear bracket sun gear Dual input/single output I I f o I1 I2 o I I f o o I1 I2 I I f o I2 o I1 I I f o f I1 I2,O I I f o f I2,O I1 I I f o f I2 I1,O I I f o f I1,O I2 Dual input/single output I I R o I1 I2 o I I R o o I1 I2 I I R o I2 o I1

I I R o f I1 I2,O I I R o f I2,O I1 I I R o f I2 I1,O I I R o f I1,O I2 Dual input/single output I I o f I1 I2 o I I o f o I1 I2 I I o f I2 o I1 I I o f f I1 I2,O I I o f f I2,O I1 I I o f f I2 I1,O I I o f f I1,O I2 Dual input/single output I I o R I1 I2 o I I o R o I1 I2 I I o R I2 o I1 I I o R f I1 I2,O I I o R f I2,O I1 I I o R f I2 I1,O I I o R f I1,O I2 Dual input/single output I o I f I1 I2 o I o I f o I1 I2 I o I f I2 o I1 I o I f f I1 I2,O I o I f f I2,O I1 I o I f f I2 I1,O I o I f f I1,O I2 Dual input/single output I o I R I1 I2 o

I o I R o I1 I2 I o I R I2 o I1 I o I R f I1 I2,O I o I R f I2,O I1 I o I R f I2 I1,O I o I R f I1,O I2 Dual input/single output I f I o I1 I2 o I f I o o I1 I2 I f I o I2 o I1 I f I o f I1 I2,O I f I o f I2,O I1 I f I o f I2 I1,O I f I o f I1,O I2 Dual input/single output I R I o I1 I2 o I R I o o I1 I2 I R I o I2 o I1 I R I o f I1 I2,O I R I o f I2,O I1 I R I o f I2 I1,O I R I o f I1,O I2 Dual input/single output f I I o I1 I2 o f I I o o I1 I2 f I I o I2 o I1 f I I o f I1 I2,O f I I o f I2,O I1 f I I o f I2 I1,O f I I o f I1,O I2

Dual input/single output R I I o I1 I2 o R I I o o I1 I2 R I I o I2 o I1 R I I o f I1 I2,O R I I o f I2,O I1 R I I o f I2 I1,O R I I o f I1,O I2

Still referring to the embodiment shown in FIG. 32, the table below entitled "Input Coupling Three Inputs" refers to an embodiment using three input torque sources to the CVT 100. For this table, reference text to CVT 100 refers to the same components as in the preceding tables, and reference text to planetary gear set 505 refers to the same components, except that "12" is indicated, which indicates components coupled to parallel paths.

transmission Input Disk(34) Idler (18) Holder(89) Output tray(101) ring gear bracket sun gear Triple input/single output I I I o I1 I2 o I I I o o I1 I2 I I I o I2 o I1 I I I o f I1 I2,O I I I o f I2,O I1 I I I o f I2 I1,O

I I I o f I1,O I2 Triple input/single output I I o I I1 I2 o I I o I o I1 I2 I I o I I2 o I1 I I o I f I1 I2,O I I o I f I2,O I1 I I o I f I2 I1,O I I o I f I1,O I2

Referring now to the embodiment shown in FIG. 34, due to the unique arrangement of the embodiments described herein, the aforementioned parallel paths can be eliminated. The parallel paths described above combine into a collinear arrangement in which the various components of the CVT and planetary gear set 505 are coupled to produce all of the combinations described above and below. In some embodiments, planetary gear set 505 is coupled to the input of CVT 100, or as shown in FIG. 34, planetary gear set 505 is coupled to the output of CVT 100. The various combinations available are listed in the table below entitled "Input Coupled Dual Output Power Paths" where there are two outputs from the CVT 100 to the planetary gear set 505. The reference text for CVT 100 is the same as in the previous table, and the reference text for planetary gear set 505 indicates the same components, except "12", which is no longer coupled to the parallel path but to the second output of CVT 100.

transmission Input Disk(34) Idler (18) Holder(89) Output tray(101) ring gear bracket sun gear

single input/dual output I o f o I1 I2 o I o f o o I1 I2 I o f o I2 o I1 I o f o f I1 I2,O I o f o f I2,O I1 I o f o f I2 I1,O I o f o f I1,O I2 single input/dual output I o R o I1 I2 o I o R o o I1 I2 I o R o I2 o I1 I o R o f I1 I2,O I o R o f I2,O I1 I o R o f I2 I1,O I o R o f I1,O I2 single input/dual output I R o o I1 I2 o I R o o o I1 I2 I R o o I2 o I1 I R o o f I1 12, O I R o o f I2,O I1 I R o o f I2 I1,O I R o o f I1,O I2 single input/dual output I f o o I1 I2 o I f o o o I1 I2 I f o o I2 o I1 I f o o f I1 I2,O I f o o f I2,O I1

I f o o f I2 I1,O I f o o f I1,O I2 single input/dual output I o o f I1 I2 o I o o f o I1 I2 I o o f I2 o I1 I o o f f I1 I2,O I o o f f I2,O I1 I o o f f I2 I1,O I o o f f I1,O I2 single input/dual output I o o R I1 I2 o I o o R o I1 I2 I o o R I2 o I1 I o o R f I1 I2,O I o o R f I2,O I1 I o o R f I2 I1,O I o o R f I1,O I2 single input/dual output f I o o I1 I2 o f I o o o I1 I2 f I o o I2 o I1 f I o o f I1 I2,O f I o o f I2,O I1 f I o o f I2 I1,O f I o o f I1,O I2 single input/dual output R I o o I1 I2 o R I o o o I1 I2 R I o o I2 o I1

R I o o f I1 I2,O R I o o f I2,O I1 R I o o f I2 I1,O R I o o f I1,O I2 single input/dual output f o I o I1 I2 o f o I o o I1 I2 f o I o I2 o I1 f o I o f I1 I2,O f o I o f I2,O I1 f o I o f I2 I1,O f o I o f I1,O I2 single input/dual output R o I o I1 I2 o R o I o o I1 I2 R o I o I2 o I1 R o I o f I1 I2,O R o I o f I2,O I1 R o I o f I2 I1,O R o I o f I1,O I2

For the two preceding tables, the transmissions described can be reversed to give inverted results for each combination, but these inverted combinations are readily recognizable and are not described separately here to save space. For example, for an output coupled dual output, input coupled/dual input inversion, it should be noted that the input of the planetary gear 505 can be coupled to any CVT 100 output.

Still referring to the embodiment shown in FIG. 34, the table below entitled "Input Coupling Dual-Dual" presents the various combinations available.

transmission Input Disk(34) Idler (18) Holder(89) Output tray(101) ring gear bracket sun gear Dual Input/Dual Output I I o o I1 I2 o I I o o o I1 I2 I I o o I2 o I1 I I o o f I1 I2,O I I o o f I2,O I1 I I o o f I2 I1,O I I o o f I1,O I2 Dual Input/Dual Output I o I o I1 I2 o I o I o o I1 I2 I o I o I2 o I1 I o I o f I1 I2,O I o I o f I2,O I1 I o I o f I2 I1,O I o I o f I1,O I2

Still referring to FIG. 34 , the table below entitled "Internal Coupled Planetary Gears in Output" presents many, if not all, of the combinations that result when planetary gear set 505 is directly coupled to components of CVT 100. With respect to CVT 100, reference text "O1" refers to the component coupled to "I1" of planetary gear set 505, "R" designates a free-rolling component or second input, and "F" designates a rigidly attached stabilizer component ( For example, a stationary housing) or a component rigidly attached to the transmission support structure, and “O2” is coupled to “I2” of the planetary gear set 505 . For planetary gear set 505, "I1" refers to the component coupled to the first output component of CVT 100, "O" refers to the component providing an output to the vehicle or other load equipment, "F" is fixed, and "I2" A second output assembly coupled to CVT 100. It should be appreciated that for the combinations described in the tables below, the input element can also be coupled to any one of the planetary elements with corresponding changes in coupling means relative to the other elements.

Input Disk(34) Idler (18) Holder(89) Output tray(101) ring gear bracket sun gear single input/dual output I O1 f O2 I1 I2 o I O2 f O1 I2 I1 o I O1 f O2 o I1 I2 I O2 f O1 o I2 I1 I O1 f O2 I2 o I1 I O2 f O1 I1 o I2 I O1 f O2 f I1 I2,O I O2 f O1 f I2,O I1 I O1 f O2 f I2,O I1 I O2 f O1 f I1 I2,O I O1 f O2 f I2 I1,O

I O2 f O1 f I1,O I2 I O1 f O2 f I1,O I2 I O2 f O1 f I2 I1,O single input/dual output I O1 R O2 I1 I2 o I O2 R O1 I2 I1 o I O1 R O2 o I1 I2 I O2 R O1 o I2 I1 I O1 R O2 I2 o I1 I O2 R O1 I1 o I2 I O1 R O2 f I1 I2,O I O2 R O1 f I2,O I1 I O1 R O2 f I2,O I1 I O2 R O1 f I1 I2,O I O1 R O2 f I2 I1,O I O2 R O1 f I1,O I2 I O1 R O2 f I1,O I2 I O2 R O1 f I2 I1,O single input/dual output I O1 O2 f I1 I2 o I O2 O1 f I2 I1 o I O1 O2 f o I1 I2 I O2 O1 f o I2 I1 I O1 O2 f I2 o I1 I O2 O1 f I1 o I2 I O1 O2 f f I1 I2,O I O2 O1 f f I2,O I1 I O1 O2 f f I2,O I1 I O2 O1 f f I1 I2,O I O1 O2 f f I2 I1,O

I O2 O1 f f I1,O I2 I O1 O2 f f I1,O I2 I O2 O1 f f I2 I1,O single input/dual output I O1 O2 R I1 I2 o I O2 O1 R I2 I1 o I O1 O2 R o 11 I2 I O2 O1 R o I2 I1 I O1 O2 R I2 o I1 I O2 O1 R I1 o I2 I O1 O2 R f I1 I2,O I O2 O1 R f I2,O I1 I O1 O2 R f I2,O I1 I O2 O1 R f I1 I2,O I O1 O2 R f I2 I1,O I O2 O1 R f I1,O I2 I O1 O2 R f I1,O I2 I O2 O1 R f I2 I1,O single input/dual output O1 O2 I f I1 I2 o O2 O1 I f I2 I1 o O1 O2 I f o I1 I2 O2 O1 I f o I2 I1 O1 O2 I f I2 o I1 O2 O1 I f I1 o I2 O1 O2 I f f I1 I2,O O2 O1 I f f I2,O I1 O1 O2 I f f I2,O I1 O2 O1 I f f I1 I2,O O1 O2 I f f I2 I1,O

O2 O1 I f f I1,O I2 O1 O2 I f f I1,O I2 O2 O1 I f f 12 I1,O single input/dual output O1 O2 I R I1 I2 o O2 O1 I R I2 I1 o O1 O2 I R o I1 I2 O2 O1 I R o I2 I1 O1 O2 I R I2 o I1 O2 O1 I R I1 o I2 O1 O2 I R f I1 I2,O O2 O1 I R f I2,O I1 O1 O2 I R f I2,O I1 O2 O1 I R f I1 I2,O O1 O2 I R f I2 I1,O O2 O1 I R f I1,O I2 O1 O2 I R f I1,O I2 O2 O1 I R f I2 I1,O single input/dual output I f O1 O2 I1 I2 o I f O2 O1 I2 I1 o I f O1 O2 o I1 I2 I f O2 O1 o I2 I1 I f O1 O2 I2 o I1 I f O2 O1 I1 o I2 I f O1 O2 f I1 I2,O I f O2 O1 f I2,O I1 I f O1 O2 f I2,O I1 I f O2 O1 f I1 I2,O I f O1 O2 f I2 I1,O

I f O2 O1 f I1,O I2 I f O1 O2 f I1,O I2 I f O2 O1 f I2 I1,O single input/dual output I R O1 O2 I1 I2 o I R O2 O1 I2 I1 o I R O1 O2 o I1 I2 I R O2 O1 o I2 I1 I R O1 O2 I2 o I1 I R O2 O1 I1 o I2 I R O1 O2 f I1 I2,O I R O2 O1 f I2,O I1 I R O1 O2 f I2,O I1 I R O2 O1 f I1 I2,O I R O1 O2 f I2 I1,O I R O2 O1 f I1,O I2 I R O1 O2 f I1,O I2 I R O2 O1 f I2 I1,O single input/dual output f O1 I O2 I1 I2 o f O2 I O1 I2 I1 o f O1 I O2 o I1 I2 f O2 I O1 o I2 I1 f O1 I O2 I2 o I1 f O2 I O1 I1 o I2 f O1 I O2 f I1 I2,O f O2 I O1 f I2,O I1 f O1 I O2 f I2,O I1 f O2 I O1 f I1 I2,O f O1 I O2 f I2 I1,O

f O2 I O1 f I1,O I2 f O1 I O2 f I1,O I2 f O2 I O1 f I2 I1,O single input/dual output R O1 I O2 I1 I2 o R O2 I O1 I2 I1 o R O1 I O2 o I1 I2 R O2 I O1 o I2 I1 R O1 I O2 I2 o I1 R O2 I O1 I1 o I2 R O1 I O2 f I1 I2,O R O2 I O1 f I2,O I1 R O1 I O2 f I2,O I1 R O2 I O1 f I1 I2,O R O1 I O2 f I2 I1,O R O2 I O1 f I1,O I2 R O1 I O2 f I1,O I2 R O2 I O1 f I2 I1,O

Figure 35 is a perspective view depicting an embodiment of the transmission 100 incorporating a planetary gear set 505 in an output coupling. In this output coupling, the parallel paths are eliminated and the planetary gear set 505 is coupled with one or more torque input sources. Planetary gear set 505 has one or two outputs coupled to respective one or two components of CVT 100. For example, in one arrangement, ring gear 524 is rigidly attached to housing 40 (not shown), and a plurality of planet gears 522 are operably attached to input disc 34 by their planet shafts 523, connected to planet shafts 523 The output of is coupled to a planetary gear carrier (not shown). In this arrangement the planet gears 522 rotate the sun gear 520 which is also attached to the holder shaft 521 which rotates the holder 89 (not shown). Each time the sun gear 520 rotates the planet gears 522 orbiting the sun gear 520 , the sun gear 520 is further rotated by the planet gears 522 rotating about their respective axes 523 . Therefore, the sun gear 520 and the retainer 89 (not shown) rotate faster than the planet gear carrier (not shown) and the input plate 34 .

Since the retainer 89 rotates faster than the input disc 34 in this configuration, the ball 1 rotates in a reversed input direction, and the direction of the shift assembly for the speed range of the CVT 100 is reversed; The direction for high speed here provides high speed, and the direction for high speed provides low speed here. When the idler gear 18 (not shown) moves toward the input side of the CVT 100, the output speed can be reduced to zero and the output disc 101 will not rotate. In other words, this occurs when the transmission is fully engaged with the rotating input but the output is not rotating. This situation is obtained by adjusting the number of teeth of the planetary gear 522 and the sun gear 520 . For example, if sun gear 520 is twice the size of planet gears 522 , sun gear 520 and retainer 89 rotate at twice the speed of planet gear carrier and input plate 34 . By increasing the speed of retainer 89 relative to the speed of input disc 34, a range in which output disc 101 rotates in opposite directions at one end of the CVT 100 switching range and a speed of output disc 101 somewhere between one end and the midpoint of the CVT 100 switching range can be produced. range of zero. The point at which the speed of the output plate 101 is zero in the switching range of the CVT 100 can be determined by dividing the speed of the sun gear 520 into the speed of the planet gear carrier, assuming that all other factors that determine the switching range providing zero output speed are equal. speed to describe.

Most, if not all, of the combinations that can be obtained by varying the embodiment shown in FIG. 35 are shown in the table below entitled "Internal Planetary Gears in Input". For reference to the components of the planetary gear set 505, "I1" refers to the output assembly coupled to the first input "I1" of the CVT 100, "I2" refers to the second output assembly coupled to the second input assembly "I2" of the CVT 100, "F" refers to the fixed assembly for planetary gear assembly 505 and CVT 100. For CVT 100, "R" refers to the freely rotating component or secondary output of torque. In this and previous tables, only the planetary ring gears are shown as fixed, any planetary element can be a fixed element, which results in more combinations. These additional combinations are not shown here to save space. Also, in the table below, only one input from the prime mover is shown. This configuration has the capability to accept two independent inputs through the planetary gears in a parallel hybrid locomotive, but these combinations are not described separately for the sake of space, and it should be understood that one of ordinary skill in the art can learn from the shown These additional combinations are understood from the examples and these descriptions. It should also be noted that any of the configurations in the table below can be combined with any of the configurations in the table above (using either a single chamber CVT or a dual chamber CVT) to produce a set of configuration.

transmission ring gear bracket sun gear Input Disk(34) Idler (18) Holder(89) Output tray(101) Dual input/single output I1 I2 IT I1 I2 f o I1 I2 IT I2 I1 f o IT I1 I2 I1 I2 f o IT I1 I2 I2 I1 f o I2 IT I1 I1 I2 f o I1 IT I2 I2 I1 f o f I1 I2, IT I1 I2 f o f I2 I2, IT I2 I1 f o f I2, IT I1 I1 I2 f o f I2, IT I2 I2 I1 f o f I2 I1, IT I1 I2 f o f I2 I1, IT I2 I1 f o f I1, IT I2 I1 I2 f o f I1, IT I2 I2 I1 f o Dual input/single output I1 I2 IT I1 I2 R o I1 I2 IT I2 I1 R o IT I1 I2 I1 I2 R o

IT I1 I2 I2 I1 R o I2 IT I1 I1 I2 R o I1 IT I2 I2 I1 R o f I1 I2, IT I1 I2 R o f I2 I2, IT I2 I1 R o f I2, IT I1 I1 I2 R o f I2, IT I2 I2 I1 R o f I2 I1, IT I1 I2 R o f I2 I1, IT I2 I1 R o f I1, IT I2 I1 I2 R o f I1, IT I2 I2 I1 R o Dual input/single output I1 I2 IT I1 I2 o f I1 I2 IT I2 I1 o f IT I1 I2 I1 I2 o f IT I1 I2 I2 I1 o f I2 IT I1 I1 I2 o f I1 IT I2 I2 I1 o f f I1 I2, IT I1 I2 o f f I2 I2, IT I2 I1 o f f I2, IT I1 I1 I2 o f f I2, IT I2 I2 I1 o f f I2 I1, IT I1 I2 o f f I2 I1, IT I2 I1 o f f I1, IT I2 I1 I2 o f f I1, IT I2 I2 I1 o f Dual input/single output I1 I2 IT I1 I2 o R I1 I2 IT I2 I1 o R IT I1 I2 I1 I2 o R

IT I1 I2 I2 I1 o R I2 IT I1 I1 I2 o R I1 IT I2 I2 I1 o R f I1 I2, IT I1 I2 o R f I2 I2, IT I2 I1 o R f I2, IT I1 I1 I2 o R f I2, IT I2 I2 I1 o R f I2 I1, IT I1 I2 o R f I2 I1, IT I2 I1 o R f I1, IT I2 I1 I2 o R f I1, IT I2 I2 I1 o R Dual input/single output I1 I2 IT I1 o I2 f I1 I2 IT I2 o I1 f IT I1 I2 I1 o I2 f IT I1 I2 I2 o I1 f I2 IT I1 I1 o I2 f I1 IT I2 I2 o I1 f f I1 I2, IT I1 o I2 f f I2 I2, IT I2 o I1 f f I2, IT I1 I1 o I2 f f I2, IT I2 I2 o I1 f f I2 I1, IT I1 o I2 f f I2 I1, IT I2 o I1 f f I1, IT I2 I1 o I2 f f I1, IT I2 I2 o I1 f Dual input/single output I1 I2 IT I1 o I2 R I1 I2 IT I2 o I1 R IT I1 I2 I1 o I2 R

IT I1 I2 I2 o I1 R I2 IT I1 I1 o I2 R I1 IT I2 I2 o I1 R f I1 I2, IT I1 o I2 R f I2 I2, IT I2 o I1 R f I2, IT I1 I1 o I2 R f I2, IT I2 I2 o I1 R f I2 I1, IT I1 o I2 R f I2 I1, IT I2 o I1 R f I1, IT I2 I1 o I2 R f I1, IT I2 I2 o I1 R Dual input/single output I1 I2 IT I1 f I2 o I1 I2 IT I2 f I1 o IT I1 I2 I1 f I2 o IT I1 I2 I2 f I1 o I2 IT I1 I1 f I2 o I1 IT I2 I2 f I1 o f I1 I2, IT I1 f I2 o f I2 I2, IT I2 f I1 o f I2, IT I1 I1 f I2 o f I2, IT I2 I2 f I1 o f I2 I1, IT I1 f I2 o f I2 I1, IT I2 f I1 o f I1, IT I2 I1 f I2 o f I1, IT I2 I2 f I1 o Dual input/single output I1 I2 IT I1 R I2 o I1 I2 IT I2 R I1 o IT I1 I2 I1 R I2 o

IT I1 I2 I2 R I1 o I2 IT I1 I1 R I2 o I1 IT I2 I2 R I1 o f I1 I2, IT I1 R I2 o f I2 I2, IT I2 R I1 o f I2, IT I1 I1 R I2 o f I2, IT I2 I2 R I1 o f I2 I1, IT I1 R I2 o f I2 I1, IT I2 R I1 o f I1, IT I2 I1 R I2 o f I1, IT I2 I2 R I1 o Dual input/single output I1 I2 IT f I1 I2 o I1 I2 IT f I2 I1 o IT I1 I2 f I1 I2 o IT I1 I2 f I2 I1 o I2 IT I1 f I1 I2 o I1 IT I2 f I2 I1 o f I1 I2, IT f I1 I2 o f I2 I2, IT f I2 I1 o f I2, IT I1 f I1 I2 o f I2, IT I2 f I2 I1 o f I2 I1, IT f I1 I2 o f I2 I1, IT f I2 I1 o f I1, IT I2 f I1 I2 o f I1, IT I2 f I2 I1 o Dual input/single output I1 I2 IT R I1 I2 o I1 I2 IT R I2 I1 o IT I1 I2 R I1 I2 o

IT I1 I2 R I2 I1 o I2 IT I1 R I1 I2 o I1 IT I2 R I2 I1 o f I1 I2, IT R I1 I2 o f I2 I2, IT R I2 I1 o f I2, IT I1 R I1 I2 o f I2, IT I2 R I2 I1 o f I2 I1, IT R I1 I2 o f I2 I1, IT R I2 I1 o f I1, IT I2 R I1 I2 o f I1, IT I2 R I2 I1 o

In the table above, it is assumed that only one CVT 100 and one planetary gear set 505 are used. It is known in the art that more planetary gear sets can produce more combinations. Since the CVT 100 described in the table can be implemented in a manner similar to a planetary gear set, it is easy for those skilled in the art to combine planetary gear sets at the input and output of the CVT 100 to establish Many more combinations, which are well known in the art, need not be listed here. However, these combinations are well within the scope of those of ordinary skill in the art and are considered part of this description.

Example

Each of the above variants has beneficial properties for specific applications. These variations can be modified and manipulated as necessary to obtain beneficial properties in any particular application. Specific embodiments will now be described using some of the variations described herein and/or listed in the above Tables. Figures 36a, 36b, 36c depict an embodiment of a transmission 3600 (which is a variation) having one source of torque input and using two sources of torque output. As before, only the main differences between the embodiment shown in Figures 36a, 36b, 36c and the previously shown and described embodiment are described. Furthermore, the described components are presented for the purpose of showing one of ordinary skill in the art how to provide power paths and torque output sources not previously described. It should be understood that many additional components can and will be used for an operational embodiment, however, to simplify the drawing, many of the above described components have been omitted and are shown schematically as boxes.

Referring to Figure 36a, torque is input through drive shaft 3669 as in the previously described embodiments. The drive shaft 3669 of this embodiment is a hollow shaft that has two ends and engages the first end with the planet gear carrier 3630 at the second end whether or not the prime mover is providing torque to the transmission 3600 . Planet gear carrier 3630 is a disc disposed coaxially with the longitudinal axis of transmission 3600 , engages drive shaft 3669 at its center and extends radially to a radius near the inside of housing 3640 of transmission 3600 . In this embodiment, the shell 3640 is stable and fixed to some support structure of the vehicle or equipment upon which it is used. Radial carrier bearings 3631 are located between the inner surface of housing 3640 and the outer edge of planet gear carrier 3630 . Carrier bearing 3631 of some embodiments is a radial bearing that provides radial support to planet gear carrier 3630 . In other embodiments, the carrier bearing 3631 is a combination ball bearing that provides radial and axial support to the planet gear carrier, preventing tilting, radial and axial movement.

A plurality of planet shafts 3632 extend from the planet gear carrier 3630 at radial locations between the center and the outer edge of the planet gear carrier 3630 . The planet shafts 3632 extend axially towards the outer end of the transmission 3600 and are generally cylindrical shafts that connect the planet gear carrier 3630 to the input disc 3634 and each of which forms a shaft about which each planet gear 3635 rotates. axis. The planet shafts 3632 may be formed on the input side of the input disc 3634 or the planet gear carrier 3630, or may be threadedly engaged to the input disc 3634 or the planet gear carrier 3630, or may be attached by a fastener, or vice versa. The planet gears 3635 are simple rotating gears that are supported by and rotate around the planet shafts 3632 , many embodiments use bearings between the planet gears 3635 and the planet shafts 3632 . These gears have straight teeth or helical teeth, however, when using helical teeth, thrust bearings are used to absorb the axial thrust generated by the transmission of the torque generated by the planetary gears 3635.

Still referring to the embodiment shown in Figure 36a, the planetary gears 3635 engage at any one time two regions along their respective circumferences as they rotate about their respective axes. At a first circumferential position furthest from the longitudinal axis of transmission 36 , each planet gear 3635 engages with ring gear 3637 . Ring gear 3637 is an internal gear formed in housing 3640 or attached to the inner surface of housing 3640 . In some embodiments, the ring gear 3637 is a set of radial teeth formed in the inner surface of the ring gear 3637 and extending radially inward such that the planet gears 3635 engage the teeth of the ring gear 3637 and are They attach along the inner surface of ring gear 3637 as they orbit along the longitudinal axis of transmission 3600 . The ring gear 3635 engages the sun gear 3620 at a point on the circumference of the planet gear 3635 generally relative to the radially outward main portion. The sun gear 3620 is a radial gear mounted coaxially about the longitudinal axis of the transmission 3600 at the center of the planet gears 3635 and engages all of the planet gears 3635 . When the planetary gear carrier 3630 revolves around the sun gear 3620 so that the planetary gears 3635 rotate, the planetary gears 3635 rotate around their respective planetary shafts 3632 through their engagement with the ring gear 3637, so the planetary gears 3635 both orbit the sun gear 3620 and in They rotate about their respective axes as they orbit. This creates rotational energy that is transmitted to the sun gear 3620 (at a speed greater than that of the drive shaft 3669 input).

In the embodiment shown in FIG. 36a , the drive shaft 3669 also drives the input disc 3634 through the planet gear carrier 3630 and the planet shafts 3632 . However, the planet gears 3635 also drive the sun gear 3620 such that power from the planet gear carrier is distributed to the input disc 3634 and the sun gear 3620 . The sun gear 3620 is rigidly connected to and rotates the holder 3689 of this embodiment. Retainer 3689 is similar to the embodiments described above, therefore, not all components are described here to simplify the drawing and improve understanding of its description. As with the other embodiments, the retainer 3689 disposes the ball 3601 about the longitudinal axis of the transmission 3600, as the retainer 3689 of this embodiment rotates about its axis, thereby orbiting the ball 3601 along the longitudinal axis of the transmission 3600. Input disc 3634 (similar to the input disc described above) provides input torque to ball 3601 in the same manner as the previous embodiment. However, sun gear 3620 also provides input torque to ball 3601 by rotating retainer 3689 and adding that torque to the input from input disc 3634 . In this embodiment, output disc 3611 is rigidly fixed to housing 3640 and does not rotate about its axis. Thus, the balls 3601 roll along the surface of the output disc 3611 as they orbit along the longitudinal axis of the transmission 3600 and rotate about their respective axes.

In other embodiments, the ball 3601 rotates the idler 3618 about its axis, however, in this embodiment the idler 3618 includes an idler shaft 3610 that extends beyond the entire portion formed by the inner diameter of the output disc 3611 . Ball 3601 drives idler gear 3618, which in turn drives idler shaft 3610, which causes transmission 3600 to output a first torque. As shown in Figure 36b, the idler shaft 3610 has a cross-sectional shape that lends itself to easy engagement with equipment that draws power from the idler shaft 3610, in some embodiments (as shown) these shapes are hexagonal, although they may Use any other shape. It should be noted that idler shaft 3610 moves axially during shifting of transmission 3600 due to the axial movement of idler gear 3618 during shifting as described below. This means that the design of the coupling between the idler shaft 3610 and the output device (not shown) allows the idler shaft 3618 to move axially. This can also be achieved by allowing the use of a slightly larger output device shaft so that the idler shaft 3610 moves freely within the output device, or by using a splined output idler shaft 3610 (eg, via ball splines). Alternatively, idler gear 3618 may be splined to idler shaft 3610 to maintain the axial position of idler shaft 3610.

Still referring to Figures 36a and 36b, the holder 3689 can also provide a source of output power. As shown, retainer 3689 is connected to retainer shaft 3690 on the inner diameter on its output side. In the illustrated embodiment, the holder shaft 3690 is formed with an output gear or spline at its end to engage and supply power as a second output source.

As shown in Figure 36a, different bearings can be implemented to maintain the axial and radial positions of the various components in the transmission 3600. The retainer 3689 is held in place by the retainer output bearing 3691, which is any radial bearing to provide radial support, or preferably a combination bearing, to maintain the axial and axial orientation of said retainer relative to the housing 3640. radial position. The holder output bearing 3691 is supported by the holder input bearing 3692 which is also radial, or preferably a radial push bearing, and sets the position of the holder 3689 relative to the input disc 3634 . In embodiments using an axial force generator (where the input disc 3634 has slight axial movement or deformation), the keeper input bearing 3692 is designed to allow such movement by any mechanism known in the industry. One embodiment uses an outer bearing race that is splined (eg, by a ball spline) to the inner diameter of the input disc 3634 in order to enable the input disc 3634 to be slightly axial relative to the outer bearing race of the retainer input bearing 3692. to move.

The switching arrangement of the embodiment shown in FIG. 36 a is slightly modified from that described to allow output torque to be supplied through idler 3618 . In this embodiment, the idler 3618 initiates switching by moving axially following actuation of the switching lever 3671 and then axially moving the switching guide 3613 (causing the switching to change the axis of the ball 3601, as described above). In this embodiment, the shift lever 3671 is not threaded into the idler 3618, but is only in contact with the idler 3618 through the idler input bearing 3674 and the idler output bearing 3673. Idler input bearing 3674 and idler output bearing 3673 are combined thrust and radial bearings, respectively, that position idler 3618 radially and axially along the longitudinal axis of transmission 3600 .

As the shift lever 3671 moves radially towards the output, the input idler bearing 3674 applies an axial force to the idler, thereby moving the idler axially to the output and causing the transmission ratio to change. The shift rod 3671 in the illustrated embodiment extends beyond the idler gear 3618 through the inner diameter formed in the center of the sun gear 3620 and into the second end of the drive shaft 3669 where the shift rod 3671 passes Idler end bearing 3675 is held within drive shaft 3669 in radial alignment. However, the shift lever 3671 moves axially within the drive shaft 3669, therefore, the idler end bearing 3675 of many embodiments allows for these movements. As previously noted, many of the above-described embodiments utilize a splined outer race that engages engaging splines formed on the inner surface of the drive shaft 3669 . These splined races allow the races to slide along the inner surface of the drive shaft 3669 as the toggle lever 3671 moves axially back and forth, and also provide radial support to help align the toggle lever 3671 radially. The inner bore of the sun gear 3620 may also be radially supported relative to the switch lever 3671 by a bearing (not shown) between the switch lever 3671 and the sun gear 3620 . The inner or outer race can again be splined to allow axial movement of the toggle lever 3671.

As the idler wheel 3618 of the embodiment shown in FIG. 36a moves axially to shift the transmission 3600 , the idler wheel 3618 moves the shift guide 3613 . In the illustrated embodiment, the switching guide 3613 is an annular ring coaxially mounted around each end of the idler gear 3618 . The shifting guide 3613 shown is held in radial and axial position by an inner shifting guide bearing 3613 and an outer shifting guide bearing 3614 . The inner and outer switch guide bearings of this embodiment are combined bearings that provide axial and radial support for the switch guide device 3613 to maintain its axial and radial alignment relative to the idler gear 3618 . Each of the shift guides 3613 may have a tubular sleeve (not shown) extending away from the idler wheel 3618 such that the shift guide bearings 3617 and 3672 may be further separated to provide additional support to the shift guide 3613 if desired . The toggle rod 3671 can be moved axially by any known mechanism for producing axial movement (eg, a vertex screw end acting as a lead screw or a hydraulically actuated piston or other known mechanism).

Referring to Figures 36a and b, and primarily to Figure 36c, the power path through the transmission 3600 follows parallel and coaxial paths. Initially, power enters transmission 3600 via drive shaft 3669 . Next, the incoming power is split and transmitted through planet gear carrier 3630 to input disc 3634 and sun gear 3620 via planet gears 3635 . The latter power path is then transmitted from the sun gear 3620 to the holder 3689 and out of the transmission 3600 via the holder shaft 3689 . Based on the size of the sun gear 3620 and the planet gears 3635, this power path provides a fixed gear ratio from the aforementioned drive shaft. The second power path is from planet carrier 3630 through planet shaft 3632 to input disc 3634 . The power path continues from input disc 3634 to ball 3601 and from ball 3601 to idler shaft 3618 and out of transmission 3600 through idler shaft 3610 . This unique arrangement allows the two power paths to be transmitted through the transmission 3600 not only in parallel paths but also coaxial paths. This type of power transmission allows the use of smaller cross-sectional sizes for the same torque transmission, with a significant reduction in volume and weight compared to other IVTs.

The embodiment shown in Figures 36a, b and c shows to those skilled in the art how the idler gear 3618 can be used as the power output listed in the table above, and how to combine the planetary gear set with the CVT Assemble as above. It is contemplated that variations of this design can be used while obtaining various desired combinations, as the number of available combinations listed is so large that such alternative designs cannot all be shown here. It should also be understood that the axial force generators presented herein could also be used with this embodiment, but for simplicity these devices are not depicted. For embodiments using one or the other of the axial force generators described herein, axial force generation can be achieved at the point where the planetary shaft 3632 is connected to the input disc 3634, although other structural arrangements may also be used. tor component is desired. In the embodiment described above, the parallel paths shown in Figures 32 and 33 are moved coaxially with the axis of the transmission 3600, which results in a very small transmission 3600 with the same torque and thereby reduces The weight and volume of this embodiment. Figure 36a, b and c output a combination to show how rotational power can be obtained from different components of different embodiments. Obviously, those of ordinary skill in the art can easily understand how to obtain other configuration structures given herein by changing the connections. In order to simplify the description of the shown combinations, it is not necessary to describe all or more combinations . The embodiments shown in Figures 35 and 36a can therefore be modified as necessary to produce any of the variations listed above or below without the need for separate non-coaxial parallel power paths.

Referring now to Figure 37a, an alternative embodiment of a transmission 3700 is shown. In this embodiment, the output disc 3711 is formed as part of the shell of the previous embodiments to form a rotating hubshell 3740 . Such an embodiment is well suited for applications such as motorcycles or bicycles. As mentioned before, only the considerable differences between this embodiment and the preceding embodiments are described in order to reduce the length of the description. In this embodiment, the input torque is applied to the input pulley 3730, which may be a pulley for a drive belt or a sprocket for a chain or some similar device. Next, the input wheel 3730 is attached to the exterior of the hollow drive shaft 3769 by press fit or with a spline coupling or other suitable method of maintaining the angular alignment of the two rotating components. Drive shaft 3769 passes through the moveable end of hub shell 3740 (this end is referred to as end cap 3741 ). The end cap is typically an annular disk having a bore through its center for drive shaft 3769 to enter the inside of transmission 3700 and having an outer diameter that mates with the inner diameter of hub shell 3740 . End cap 3741 can be secured to end cap 3740 or threaded into the hub shell to encapsulate the internal components of transmission 3700 . The end cap 3741 of the illustrated embodiment has a bearing surface and corresponding bearing on the inside of its outer diameter for positioning and supporting the axial force generator 3760, and a bearing surface and corresponding bearing on its inner diameter. A bearing is supported between the end cap 3741 and the drive shaft 3769.

Drive shaft 3769 is mounted on and rotates about input axle 3751 , which is a hollow tube secured to vehicle frame 3715 by frame nut 3752 and provides support for transmission 3700 . The input axle 3751 incorporates a shift lever 3771 similar to that described in previous embodiments (such as that shown in FIG. 1 ). The shift lever 3771 of this embodiment is actuated by a shift cap 3743 that is threaded onto the end of the input axle 3751 extending beyond the vehicle frame 3715 . The shift cap 3743 is a tubular cap having a set of internal threads formed on the inner surface of the cap and capable of mating with a complementary set of external threads formed on the outer surface of the input axle 3751 . The end of the switch rod 3771 extends through a hole formed in the input end of the switch cap 3743 and is itself threaded, which allows the switch cap 3743 to be secured to the switch rod 3771. By rotating the switch lever 3771, its thread (which can be acme threading or any other thread) makes it move axially, because the switch lever 3771 is fixed to the switch cap 3743, so the switch lever 3771 is also axially movement, which causes shifting guide 3713 and idler 3718 to shift, thereby shifting transmission 3700.

Still referring to the embodiment shown in FIG. 37a, the drive shaft 3769 relies on and is supported by the input shaft 3751 and one or more shaft support bearings 3772, which may be needle bearings or other radial support bearings. The drive shaft 3769 applies torque to the axial force generator 3760, as shown in the previous embodiments. Any of the axial force generators described herein can be used with transmission 3700, and this embodiment utilizes a screw 3735 that is driven by drive shaft 3769 through splines or other suitable distribution of torque to drive plate 3734 and The bearing disc 3760 is driven by means of the bearing disc 3760, as in the previous embodiment. In this embodiment, a drive seal 3722 is disposed between the inner diameter of the input wheel 3770 and the outer diameter of the input wheel shaft 3751 (beyond the end of the drive shaft 3769) to limit the amount of foreign matter allowed to enter the interior of the transmission 3700 . Other seals (not shown) may also be used between the housing cap 3742 and the input wheel to limit foreign particle penetration from between the end cap 3741 and the drive shaft 3769 . Drive seal 3722 may be an o-ring seal, lip seal, or other suitable seal. This illustrated embodiment also utilizes a similar retainer 3789 as the previous embodiments, however, the illustrated transmissions 3700 use axial bearings 3799 to support the balls 1 in their axle 3703 . The bearings 3799 may be needle bearings or other suitable bearings and reduce friction between the balls and their axles 3703 . Any of the various embodiments described herein (or those known to those skilled in the art) balls and axles may be used to reduce the resulting friction.

Still referring to the embodiment shown in FIG. 37 a , the holder 3789 and shift lever 3771 are supported by the output axle 3753 in the aforementioned output side. The output axle 3753 is a somewhat tubular member that sits in a bore formed in the output end of the hub shell 3740, between the retainer 3789 and the vehicle frame 3715 on the output side. The output shaft 3753 has a bearing race and bearings formed between its outer diameter and the inner diameter of the hub shell 3740 to allow relative rotation of the two components as the output shaft 3753 supports the output side of the transmission 3700 . The output shaft is clamped to the vehicle frame 3715 by an output support nut 3754.

As shown in FIG. 37 a , the actuator 3700 is switched by applying tension to a switch cable 3755 that is coiled and applies a rotational force to the switch cap 3743 . Switching cable 3755 is a tether capable of applying tension and actuated by a switching device (not shown) used by an operator to switch transmission 3700 . In some embodiments, the switching cable 3755 is a guide wire capable of pulling and pushing such that only one coaxial guide wire (not shown) is required to run from the transmission 3700 to the switching device described above. Depending on the switching device the operator is using, the switching cable 3755 is routed through the frame bezel 3716 to and out of the aforementioned switching cap. Frame fence 3716 is a portion extending from vehicle frame 3715 that guides switch cable 3755 to switch cap 3743 . In the illustrated embodiment, the baffle guides 3716 are somewhat cylindrical extensions that have slots formed along their length through which the switch cables 3755 are passed and guided. In another aspect, the transmission 3700 shown in Figure 37a is similar to other embodiments shown herein.

An embodiment similar to that shown in Figure 37a is depicted in Figure 37b. In this embodiment, the output disk 3711 is also fixed to the housing 3740, however, the housing 3740 is fixed and cannot rotate. However, in this embodiment, the retainer 3789 is free to rotate relative to the output disc 3711 and housing 3740, similar to the embodiment shown in FIG. 36a. This means that the output goes through idler 3718 again. In this embodiment, an idler 3718 is attached to a movable output shaft 3753 similar to that described for the embodiment in Figure 36a. The output shaft 3753 terminates at the distal end on the output side in output splines 3754, which allow the transmission 3700 to torque-couple the movable output shaft 3753 to other devices. In this embodiment, torque is applied to transmission 3700 via output shaft 3772 by a chain and sprocket (not shown), input gear (not shown), or other known coupling means. This torque then passes through to the input disc 3734, as described in the previous embodiments. However, as described, with reference to FIG.

As with the embodiment shown in Figure 36a, the switching guide 3713 of this embodiment is supported by bearings 3717 on the outer surface of the output shaft 3753 by applying a torque output via an idler gear 3718. The transmission 3700 is switched by moving the switching lever 3771 axially and actuated by the actuator 3743 . The actuator may be the switching cap in Figure 37a, or a wheel or gear controlled by a motor or manually, or the actuator 3743 may be any other device for axially setting the position of the switching lever, such as One or more hydraulic pistons. In some embodiments, an axial force generator 3960 and switching device are used as will be shown in Figure 39a below. With these embodiments, very high transmission ratios can be obtained with very high efficiency and with very low frictional losses compared to other types of transmissions.

Figure 38 shows an alternative embodiment of a ball shaft 3803 that can be used with many of the transmissions described herein. In this embodiment, oil is injected into the bore in the ball 1 through threads 3810 formed in the outer diameter of the ball shaft 3803 . Adhering to the surface of the ball 1 adjacent to the above-mentioned hole is an oil layer, which is drawn around the ball shaft 3803 at the same speed as the surface to which it is attached when the ball 1 rotates around the ball shaft 3803 and advances; The respective distance of the layers additionally pulls the solidified adjacent oil layers with reduced cohesive strength by the same attractive force used to create the viscosity of the oil. As the aforementioned oil layer is drawn around the ball shaft, the leading edge of a particular volume of oil in the layer is sheared by the surface of a set of threads 3810 formed in the outer surface of the ball shaft 3803 . Threads 3810 may be subacimer threads or other types of threads suitable for the injection operations described herein. As each volume of oil is sheared from an adjacent layer (which is on the outside of the threads 3810), it may be replaced by a similar layer that is then sheared by the same action. Since the threads 3810 are shaped such that they are introduced into the bore of the ball 1 , the sheared volumes of oil move into the ball 1 as they are constantly replaced by the shear forces that follow them. As the above-mentioned operation continues, the oil enters the hole of the ball 1 by its own self-attraction, and a kind of suction is established. This "pumping" action is therefore proportional to the viscosity of the oil. To facilitate this pumping effect, in many embodiments it is chosen to use a lubricant that behaves as a Newtonian fluid at shear rates over the range of rotation rates of the ball 1 in any particular embodiment.

Still referring to FIG. 38 , the thread 3810 begins at a position along the axis of the ball shaft 3803 slightly outside the edge of the ball 1 in order to create an override of the shearing action of the oil flowing into the ball 1 . The distance that the threads 3810 extend to the outside of the ball 1 can be between 0.5 thousandths of an inch and 2 thousandths of an inch, while in other embodiments the distance can be 10 thousandths of an inch to 1 thousandth of an inch, or more Or less, depending on manufacturing cost and other considerations. In the illustrated embodiment, the thread 3810 extends into the bore of the ball 1 and at a location within the ball in a pocket 3820 formed by the longitudinal length of the ball shaft 3803 at a radius smaller than elsewhere in the ball shaft 3803. stop. Inside the ball 1, the container 3820 terminates at the container end 3830, where the outside diameter of the ball shaft increases again to a size close to the inside diameter of the ball 1, thereby forcing the oil through the gap between the ball shaft 3803 and the inside surface of the ball 1. A small gap leaks from the ball 1, which creates a high pressure oil supply that is used to create a lubricant film between the two components mentioned above. In some embodiments, no container 3820 is present and the threads 3810 simply terminate near the middle of the ball.

By controlling the size of the gap between the ball shaft 3803 and the inner surface of the ball 1, a balance can be achieved between the amount of leaked oil and the amount of injected oil, thereby maintaining the lubricating pressure in the hole of the ball 1. This balance depends on the viscosity of the oil, the size of the gap and the rotational speed of the ball 1 . Although the container end 3830 is shown near the middle of the ball 1, this is for descriptive purposes only and the container 3820 could terminate near the other end of the ball 1, or closer to the threads 3810, depending on the application. In other similar embodiments, this same orientation is formed by threads formed through the interior of the bore of the ball 1, similar to that shown in Figure 23, except that the threads 3810 are as in this embodiment Terminates in container 3820 formed near the middle of ball 1 and ball shaft 3803 as described in .

Referring now to Figures 39a, b and c, other alternative axial force generators 3960 are shown. In these embodiments, the screw 3935 is located in the bore of a bearing disc (not shown), rather than in the input disc 3934. In this embodiment, screw 3935 is directly driven by a drive shaft (not shown) via splines 3975 which mate with mating splines of the drive shaft. Screw 3935 in turn distributes torque to input disc 3934 via screw ramp 3998 and center disc ramp 3999, and via its threads 3976 and a corresponding set of internal threads (not shown) formed on the inner surface of the bore of the aforementioned bearing disc to the bearing disc. As the screw 3935 is rotated by the drive shaft, a set of central screw ramps 3998 formed on the output end of the screw 3935 is rotated and engages and rotates a corresponding set of central disc ramps 3999 . Center disk ramp 3999 is formed on the thrust washer surface formed near the inner diameter of the input side of input disc 3934, and as they are rotated by center screw ramp 3998, through the reaction of the angled surfaces of center 3999, center ramp 3998 and 3999 begins to apply torque and axial force to input disc 3934. In addition, rotation of the screw 3935 engages its threads 3976 with the threads of the bearing disc to begin rotating the bearing disc.

Referring now to FIG. 39 a , in the illustrated embodiment, the position of the idler 3918 directly affects the axial force generator 3960 . In this embodiment, the idler assembly has a tubular extension, referred to as a pulley land 3930 , extending from the thrust guide 3713 on the input side in a radially outward tubular extension terminating in the vicinity of the input disc 3934 . The connecting assembly consisting of fixed link 3916, first link pin 3917, short link 3912, cam bar 3914, cam link pin 3915 and fixed cam pin 3923 extends axially from pulley table 3930 toward screw rod 3935 and moves according to the transmission The position of the screw 3935 is set axially. Links 3916, 3912, and 3914 are generally elongated struts. The fixed link 3916 extends from the input end of the pulley table 3930 towards the threaded rod 3935 and is connected to the middle short link 3912 by the first link pin 3917 . The first link pin 3917 forms a floating pin joint between the fixed link 3916 and the short link 3912 so that when the two links 3916, 3912 move axially during switching, the short link 3912 Can rotate about the first link pin 3917. The short link 3912 is connected at its other end to the cam link 3914 by the cam link pin 3915, thereby forming a floating hinge joint. The cam link 3914 is axially secured by a fixed cam pin 3923 which is secured to the shaft 3971 or other fixed assembly and forms a hinge joint about which the cam link 3914 rotates as the idler gear 3918 moves axially .

In the following description, in order to simplify the drawings, the bearing disc 60, the ramp bearing 62, the outer peripheral ramp 61 and the input disc ramp 64 in FIG. 1 are not separately shown, but in this embodiment, similar components can be utilized and perform similar functions. When the axial force generator 3960 shown in Figures 39a, 39b and 39c is at a high transmission ratio, the idler gear 3918 is positioned axially away from its input side, so the fixed link 3916 is also located towards its input side furthest axial position. First link pin 3917, short link 3912, and second link pin 3921 are all positioned toward the input side, so cam link 3914 is oriented about fixed cam pin 3923 such that its cam surface (not separately depicted) is away from threaded rod 3935 rotate. The cam link 3914, as it rotates about the axis of its fixed cam pin 3923, applies a cam force to the screw 3935 to force it toward the output side at low gear ratios. However, at low gear ratios, the cam surface of the cam link 3914 rotates away from the screw 3935, as shown. This allows the screw 3935 to be settled in its furthest position towards the output side and causes the bearing coil to rotate counterclockwise around the screw 3935 (as viewed from the input side towards the output side) in order to maintain engagement with the screw threads 3976. With this rotation, the bearing ramp rotates counterclockwise to allow the aforementioned bearing disc (not shown here, but similar to the disc bearing shown in FIG. , The position where the bearing does not provide or provide a small amount of axial force.

At the same time, due to the screw rod 3935 to the extreme position to the left (as shown in Figure 39a), the center screw rod slope 3998 is fully engaged with the center disk slope 3999, thereby making the input disk 3934 rotate clockwise slightly to allow the axial direction of the screw rod 3935 position is at its furthest output side position. Rotation of the input disc 3934 in this manner means that the input disc ramp rotates in the opposite direction to the bearing disc ramp, amplifying the effect of unloading the peripheral ramp and bearing. In this case, the majority or all of the axial force is applied through the central ramps 3998, 3999, with little force applied if the axial force is generated by the aforementioned peripheral ramps.

When the idler gear 3918 moves toward the output side to shift to a low gear ratio, the linkage mechanism described above begins to extend as the fixed link 3916 moves axially away from the screw 3935 and the cam link 3914 rotates about the fixed cam pin 3923. When the cam link 3914 rotates around the fixed cam pin 3923, the axial movement of the fixed link 3916 acts on one end of the cam link 3914, while the other end moves towards the screw 3935, thus reversing the applied force of the fixed link 3916. the direction of the axial force. By adjusting the lengths of the various connections to the cam levers 3914, the axial force exerted by the fixed link 3916 can be reduced or amplified through leverage. The cam end of the cam link 3914 applies an axial force to the thrust washer 3924 on the output side of the screw 3935 . Thrust washer 3924 engages thrust washer bearing 3925 and bearing race 3926 to apply a resultant axial force to screw 3935 . In response, the screw 3935 moves axially towards the input side, and its threads 3976 cause the bearing disk to rotate clockwise (viewed from the input side to the output side), which causes the peripheral ramp to rotate, causing the ramp bearings to move along the peripheral ramp until they begin to produce The location of the axial force. Simultaneously, due to the axial movement of the screw rod 3935 towards the input side, the central screw ramp 3998 disengages from the central disc ramp 3999, and the input disc 3934 rotates counterclockwise relative to the screw rod 3935, again causing the outer peripheral ramp bearing to move to the position where the axial force is generated. Location. Through this leverage of the linkage mechanism, the axial force generator 3960 of this embodiment effectively distributes axial force and torque between the central ramps 3998, 3999 and the peripheral ramps.

Also shown in Fig. 39a is an optional leg assembly to that shown in Fig. 5 that can reduce the overall bulk of the transmission. In this depicted embodiment, the rollers 3904 are disposed radially inwardly of the legs 3902 (compared to leg 2 in FIG. 5 ). Furthermore, the input disc 34 and output disc (not shown) contact the balls 1 close to their shafts, which reduces the load on the idler 18 and allows the transmission to carry more torque. As a result of these two modifications, the overall diameter of the input disk 34 and output disk (not shown) of this embodiment is reduced to be substantially equal to the diameter at the farthest opposing point of the two directly opposing balls 3901 of this embodiment , as shown by the line "O.D".

A further feature of the embodiment shown in Figure 39a is a modified switching assembly. The rollers 3904 of this embodiment are formed as pulleys each having a concave radius 3905 (rather than a convex radius) on their outer edges. This allows the rollers 3904 to perform the function of aligning the ball shafts 3903, but also allows them to act like pulleys to change the axis of the ball shafts 3903 and balls 3901 to switch transmissions. The flexible cable 155 described with reference to Figures 1 to 6, or a similar switching cable, can be wrapped around the rollers 3904 so that when tension is applied, the rollers 3904 are brought closer together to switch the transmission. The above-mentioned switching cable (not shown in FIG. 39 ) is guided to the roller 3904 through the holder (shown at 89 in FIG. 1 ) by the guide roller 3951, which is also a pulley in this illustrated embodiment, which Installed on the guide shaft 3952 at the output end of the pulley table 3930.

In some embodiments, guide roller 3951 and guide shaft 3952 are designed to allow the axis of guide roller 3951 to pivot on a pivot in order to maintain contact with roller 3904 as the angle of ball axle 3903 relative to the axis of the transmission changes. Pulley type alignment. In some embodiments, this can be accomplished by mounting guide shaft 3952 to pulley table 3930 using a ring-pivot joint or pivot or any other known method. In this embodiment, a switch cable acts on a set of rollers 3904 in either the input or output side of the ball 3901, and a spring (not shown) biases the ball shaft 3903 to switch in the other direction. In other embodiments, two switching cables are used, one on one side that pulls the roller 3904 radially inward on that side, and the other on the opposite side of the ball 3901. The cable pulls the rollers 3904 on each side radially inward, thus switching the transmission. In such an embodiment, a second pulley table 3930 or other suitable structure may be formed on the output end of the switch guide 3913 with a corresponding set of guide shafts 3925 and guide rollers 3951 mounted on the second pulley table 3930 . Cables (not shown) in this embodiment pass through holes or slots (not shown) formed in the hub 3971 and exit the transmission via the hub 3971 . The cables can pass from either or both ends of the hub 3971, or they can pass through hubs 3971 formed axially beyond the input disc (not shown) and output disc (not shown) or bushings. (not shown) in the additional hole at any one or both ends, and the above-mentioned output disk is a rotary shaft sleeve. The holes or slots through which the cables pass are designed to maximize the life of the cable material through the use of radiused edges and pulleys, and these devices are used in various locations on axles and transmissions for conveying cables.

Referring to Figures 39a, 40a and 40b, one embodiment of a connection assembly 4000 of the axial force generator 3960 of Figure 39a is shown. The illustrated connection assembly 4000 is also comprised of fixed link 3916, first link pin 3917, short link 3912, second link pin 3921, cam link 3914, cam link pin 3915, and fixed cam pin. The stationary link 3916 of this embodiment is an elongated strut having a first end rigidly attached to the pulley table 3930 shown in FIG. Hinge joint holes are formed at the two ends. Fixed links 3916 are generally parallel to the sides of axle 3971 . The first link pin 3917 is disposed in the hole in the second end of the fixed link 3916, and combines the second end of the fixed link 3916 with the first end of the short link 3912, and the first end of the short link 3912 having corresponding pivot holes formed therein. The cam link pin 3915 is arranged in the hole of the second end of the short link 3912, and utilizes the first end of the cam link 3914 to connect with the second end of the short link 3912 via the hinge joint hole formed in the cam link 3914. end connection. The cam link 3914 has two ends, a first end and an opposing male end having a raised surface 4020 formed in its outer edge. The cam link 3914 also has a second hinge engaging hole formed in the middle between its first end and the protruding end, and the fixed cam pin 3923 passes through the second hinge engaging hole. Fixed cam pin 3923 is fixed to a fixed part of the transmission (eg, axle 3971 ) such that it forms an axis about which cam link 3914 rotates.

Figure 40a shows the linkage assembly 4000 in a retracted state, corresponding to a very high transmission ratio, where the fixed link moves in all directions towards the input of the transmission, as shown in Figure 39a. Figure 40b shows the connection assembly 4000 in an extended state, corresponding to a low transmission ratio. As described above, the cam link 3914 applies an axial force to the screw 3935 such that as the transmission switches from high to low, the axial force is switched from being generated by the central ramps 3998, 3999 to being generated by the outer peripheral ramps. Additionally, when the transmission is switched from low to high, the cam link 3914 reduces the amount of axial force applied to the screw 3935 to allow the screw 3935 to move axially towards the output end, thereby diverting the axial force from that generated by the peripheral ramp. Switch to be generated by center bevel 3998,3999.

As shown in Figures 40a and 40b, the cam surface 4020 of the cam link 3914 can be designed to provide various loading and unloading profiles. In fact, in this embodiment, the second cam surface 4010 is disposed on the first end of the cam link 3914 . As shown in Figure 40a, at very high gear ratios, the cam surface 4020 is completely unloaded, applying minimal, if any, axial force to the screw 3935. However, in some embodiments, it may be necessary to apply greater axial force at different transmission ratios, in which case, at the highest transmission ratio, the second cam surface 4010 increases the axial force to the screw , thereby transmitting some of the generated axial force back to the outer peripheral disc, thereby increasing the amount of axial force required at high transmission ratios. This is merely an example of variations that can be included to vary the control of the axial force generator 4060 to generate axial force depending on the torque-speed profile desired for a particular application.

The embodiments described herein are examples given to satisfy legal descriptive requirements. These examples are merely embodiments that can be used by either party, and they should not limit the invention in any way. Accordingly, the invention is defined by the appended claims, not by any examples or terms herein.

Claims (37)

1. A variable speed transmission comprising: longitudinal axis; a plurality of balls distributed radially about said longitudinal axis, each of said balls having a tiltable axis and rotating about said axis; a rotatable input disc positioned adjacent to and in contact with each of the balls; a rotatable output disc disposed adjacent the balls and opposite the input disc and in contact with each of the balls; a rotatable idler having a substantially constant outer diameter coaxial about said longitudinal axis and disposed radially inwardly of and in contact with each of said balls; and A planetary gear set is coaxially mounted about the longitudinal axis of the transmission. 2. The variable speed transmission of claim 1, wherein said balls aggregate torque components transmitted from at least two power paths provided by said planetary gearset, wherein said at least two The power paths are coaxial. 3. The variable speed transmission of claim 1 wherein at least one of said idler gear and said output plate provides a torque input to said planetary gearset. 4. The variable speed transmission of claim 1, wherein said idler has an irregular bore, said idler being shifted by first and second gears mounted coaxially within said bore. The guide bushing is supported at a radial location of the idler and is further supported by first and second radial support bearings located at each end of the idler bore parts and their respective guide bushings. 5. A variable speed transmission as claimed in claim 4, wherein said transmission is shifted by axially moving said guide bushing. 6. The variable speed transmission of claim 1, wherein said planetary gear set further comprises: a ring gear mounted coaxially about said longitudinal axis and having radially inwardly disposed teeth; a plurality of planet gears coaxially distributed within and engaging the ring gear about the longitudinal axis, each of the planet gears having a respective planetary axis and rotating about the planetary axis, wherein said planet axis is radially remote from said longitudinal axis; a plurality of planet gear shafts, each of said planet gears having one said planet gear shaft, said planet gears rotating about said planet gear shafts; a sun gear coaxially mounted about the longitudinal axis and radially within and engaged with each of the plurality of planet gears; and A planet carrier mounted coaxially about the longitudinal axis and adapted to support and position the planet shafts. 7. The variable speed transmission of claim 6, further comprising a retainer adapted to align said tiltable axis of said ball, said retainer further adapted to retain said ball angular and radial position. 8. The variable speed transmission of claim 7 wherein at least one of said idler gear, said retainer and said output plate provides a torque input to said planetary gearset. 9. A variable speed transmission as claimed in claim 7 wherein said planet carrier is applied with input torque and is coupled to said input disc, said sun gear is coupled to said retainer, said annular The gears are fixed and cannot rotate, and output torque is provided from the transmission through the output disc. 10. The variable speed transmission of claim 6, further comprising an axial force generator adapted to generate increased traction between said input disc, said balls, said idler and said output disc axial force. 11. The variable speed transmission of claim 10, wherein the amount of axial force generated by the axial force generator is a function of the transmission ratio of the transmission. 12. A variable speed transmission as claimed in claim 11 wherein each of said input disc, said balls, said output disc and said idler has a contact surface coated with increasing Rubbing paint. 13. The variable speed transmission of claim 12, wherein the paint is ceramic. 14. The variable speed transmission of claim 12, wherein the coating is a cermet. 15. The variable speed transmission of claim 12, wherein the paint is selected from silicon nitride, silicon carbide, electroless nickel plating, electroplated nickel or any combination thereof. 16. The variable speed transmission of claim 12, wherein the coating has a thickness between 0.25 microns and 5 microns. 17. The variable speed transmission of claim 12, wherein the coating has a thickness between 0.5 microns and 4 microns. 18. A variable speed transmission comprising: longitudinal axis; a plurality of balls distributed radially about said longitudinal axis, each of said balls having a tiltable axis and rotating about said axis; a rotatable input disc positioned adjacent to and in contact with each of the balls; a fixed output disc disposed adjacent the balls and opposite the input disc and in contact with each of the balls; a rotatable idler having a substantially constant outer diameter disposed radially inwardly of and in contact with each of said balls; a retainer adapted to retain the radial position and axial alignment of the balls, and the retainer is rotatable about the longitudinal axis; and An idler shaft, coupled to the idler, is adapted to receive a torque output from the idler and transmit the torque output from the transmission. 19. A variable speed transmission as claimed in claim 18, further comprising an axial force generator adapted to generate increased traction between said input disc, said balls, said idler and said output disc axial force. 20. The variable speed transmission of claim 19, wherein the amount of axial force generated by the axial force generator is a function of the transmission ratio of the transmission. 21. A variable speed transmission as claimed in claim 20 wherein each of said input disc, said balls, said output disc and said idler has a contact surface coated with Coatings that increase friction. 22. A variable speed transmission as claimed in claim 21 wherein said paint is ceramic. 23. The variable speed transmission of claim 21 wherein said coating is a cermet. 24. The variable speed transmission of claim 21, wherein the coating is selected from silicon nitride, silicon carbide, electroless nickel plating, electroplated nickel, or any combination thereof. 25. The variable speed transmission of claim 21, wherein the coating has a thickness between 0.25 microns and 5 microns. 26. The variable speed transmission of claim 21, wherein the coating has a thickness between 0.5 microns and 4 microns. 27. A variable speed transmission comprising: longitudinal axis; a first set of balls radially distributed about said longitudinal axis, each of said first set of balls having a tiltable axis and rotating about said axis; a second set of balls radially distributed about said longitudinal axis, each of said second set of balls having a tiltable axis and rotating about said axis; a rotatable first input disc positioned adjacent to and in contact with each of the first set of balls; a rotatable second input disc positioned adjacent to said second set of balls and in contact with each of said first set of balls; an input shaft coaxial with said longitudinal axis and connected to said first and second input discs; a rotatable output disc positioned between and in contact with each of the first and second sets of balls; a generally cylindrical first idler wheel disposed radially inwardly of and in contact with each of said first set of balls; A generally cylindrical second idler is disposed radially inwardly of and contacts each of the second set of balls. 28. A variable speed transmission as claimed in claim 27, further comprising an axial force generator adapted to generate increased traction between said input disc, said balls, said idler and said output disc axial force. 29. A variable speed transmission as claimed in claim 28, wherein the amount of axial force generated by the axial force generator is a function of the gear ratio of the transmission. 30. A variable speed transmission as claimed in claim 29 wherein each of said input disc, said balls, said output disc and said idler has a contact surface coated with Coatings that increase friction. 31. The variable speed transmission of claim 30, wherein the paint is ceramic. 32. The variable speed transmission of claim 30, wherein the coating is a cermet. 33. The variable speed transmission of claim 30, wherein the paint is selected from silicon nitride, silicon carbide, electroless nickel plating, electroplated nickel layer or any combination thereof. 34. The variable speed transmission of claim 30, wherein said coating has a thickness between 0.25 microns and 5 microns. 35. The variable speed transmission of claim 30, wherein the coating has a thickness of between 0.5 microns and 4 microns. 36. A rolling traction variable speed transmission comprising: longitudinal axis; a plurality of circular speed adjusters distributed around said longitudinal axis, each said circular speed adjuster having a tiltable shaft and rotating about said shaft; input discs coaxial with and rotatable about the longitudinal axis, the input discs disposed adjacent to and in contact with each of the circular speed adjusters; output discs coaxial with and rotatable about the longitudinal axis, the output discs disposed adjacent to and in contact with each of the circular speed adjusters; and an axial force generator adapted to apply an axial force to increase contact force between the input disc, the output disc and the plurality of speed adjusters, the axial force generator further comprising: a bearing disc coaxial with said longitudinal axis and rotatable thereabout, said bearing disc having an outer diameter and an inner diameter and having a threaded hole formed in its inner diameter; a plurality of peripheral ramps attached to a first side of the bearing disc proximate its outer diameter; a plurality of bearings adapted to engage the ramps of the plurality of bearing discs; a plurality of input disc peripheral ramps mounted on the input disc on opposite sides of the speed regulator and adapted to engage the bearing; a generally cylindrical screw coaxial with and rotatable about said longitudinal axis, said screw having male threads formed along its outer surface adapted to engage threaded holes in said bearing disc; a plurality of central screw ramps attached to the end of the screw facing the speed regulator; and A plurality of central input disc ramps attached to the input disc and adapted to engage the plurality of central screw ramps. 37. The rolling traction variable speed transmission of claim 36, wherein said axial force generator further comprises: a thrust bearing in contact with the end of said screw; a linkage disposed along said longitudinal axis and adapted to apply an axial force to said thrust bearing, said linkage tending to axially move said screw away from said input disc in response to switching of said transmission ratio , wherein when the transmission shifts, the linkage applies an appropriate amount of axial force to the thrust bearing as a function of the transmission ratio. 38. A vehicle comprising: engine; powertrain; and Variable speed transmission, including: longitudinal axis; a plurality of balls radially distributed about said longitudinal axis, each ball having a tiltable axis and rotating about said axis; a rotatable input disc positioned adjacent to and in contact with each of the balls; a rotatable output disc disposed adjacent the balls and opposite the input disc and in contact with each of the balls; a rotatable idler having a substantially constant outer diameter coaxial about said longitudinal axis and disposed radially inwardly of and in contact with each of said balls; and A planetary gear set is coaxially mounted about the longitudinal axis of the transmission.

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