GB2032540A - Friction transmission - Google Patents

Abstract

A friction transmission comprises a driving member 101 and a driven member 102 mounted for rotation about a common axis, the members having oppositely facing annular arcuate surfaces 31, 32 and at least one roller 103 contacting said surfaces 31, 32 and supported with its axis intersecting the common axis of said members 101, 102. Lines tangent to the roller 103 and the annular races at the points of contact converge at the intersection of the axis of rotation of the roller 103 and the common axis of the members 101, 102. Hydraulic means 171, 170 apply a force to the roller 103 in the direction of the roller axis proportional to the driving load. The roller 103 and roller carrier 104 are supported by trunnions 135 which are in turn supported by opposed resilient means. Means are provided for cleaning and coating the friction surfaces and a torque limiting clutch prevents slipping of the traction contacts due to transient impulses of excess torque. <IMAGE>

Description

SPECIFICATION Friction transmission This invention relates to a friction transmission with opposed toric races and interposed conical rollers.

It has heretofore been suggested that a friction transmission can be provided by utilizing coaxial driving and driven members with oppositely facing toric race surfaces in which rollers are provided in contact with the races. The axes of the rollers intersect the common axis of the driving and driven members and by tilting the rollers along the axis of the driving and driven members, a variable drive is obtained. It has heretofore been recognized that when tangents to the points of contact do not converge at the intersection of the axes of the rollers and the axis of the driving and driven members, there is slippage or spin action, resulting in wear and loss of effective tractive force.

Among the features of the invention are: 1. True rolling action to eliminate "spin" at the contact areas. The lines tangent to the contacting surfaces meet the main shaft axis at the intersection of the roller axes at all times.

2. True rolling action as between cones with a common apex so that there is no need to lubricate the contacting areas. The space in which the rollers operate may be sealed and drained to avoid deposit of lubricant on the friction surfaces. The coefficient of friction may be further increased by the selection of race and roller materials and perhaps by coatings or other means. Friction coefficients many times higher than with lubricated surfaces may be obtained. This reduces the required contact pressure and permits the use of conventional thrust bearings of reasonable size.

3. Planetary differential gears to neutralize the motion of the output shaft at a specific position (ratio) of the rollers. This provides (a) a neutral or stationary position with the transmission fully engaged and operating and (b) a gradual forward motion as the ratio is increased and (c) reverse motion as the ratio is decreased. If desired, the differential gears may be bypassed and the driven member connected to the output shaft directly at a suitable output shaft speed.

4. Control of ratio change by balancing the reactive force of the rollers in line with the tilt axis by (a) a hydraulic piston or (b) mechanical means.

Movement along the tilt axis of the roller carrier in a plane normal to the main shaft axis in one direction induces a spiral action to increase the ratio, and in the other direction to decrease it.

Change in ratio produces a change in the reactive force which moves the roller carrier axially to reduce or stop the spiral movement.

5. Control of the contact pressure on the rollers in proportion to the reactive force by hydraulic means.

6. A means for cleaning and/or coating the friction surfaces.

7. A torque limiting clutch to protect the traction contacts from slipping due to transient impulses of excess torque.

The invention provides an efficient means of matching speeds of high inertia mechanical elements over a range of ratios precisely and, if desired, automatically. Examples are: 1. Vehicles having prime movers which operate most efficiently with a rather narrow range of speeds, such as a single shaft gas turbine.

2. Vehicles with hybrid power plants, for example, an engine driving a high speed flywheel.

Such an engine, of whatever type, would only need to produce enough power to drive the vehicle at a sustained maximum speed. When the maximum power is not required, the excess energy would be stored in the flywheel and extracted as needed for acceleration, hill climbing or bursts of speed for passing. The engine could operate at maximum efficiency, perhaps intermittently. The cooling system and emission controls would be much smaller and would operate under much more uniform conditions.

Regenerative braking would return large amounts of power to the flywheel. This should greatly improve the fuel efficiency of automobiles, urban busses and trucks operating in hilly terrain.

3. Battery powered vehicles with constant speed motors and regenerative braking.

4. Flywheel storage of off-peak electric generator power.

5. Flywheel storage of wind or solar power to drive electric generators at constant speed.

In accordance with the invention, the friction transmission comprises convex conical rollers which engage opposed surfaces of a driving and driven member, which surfaces have a configuration such that lines tangent to the periphery of the rollers and the opposed race surfaces converge at the intersection of the axes of rotation of the rollers and the common axis of the members at all positions of the rollers and not requiring lubrication, means for applying and removing cleaning and/or conditioning material to remove contamination from the friction surfaces, and applying coating material to maintain or increase the coefficient of friction, and means to prevent slipping of the traction contact surfaces due to excess torque.

The present invention will now be described, by way of example, with reference to the accompanying drawings in which: Fig. 1 is a longitudinal sectional view of a friction transmission embodying the invention.

Fig. 2 is a diagrammatic view of a portion of the transmission shown in Fig. 1.

Fig. 3 is a sectional view taken along the line 3-3 in Fig. 1.

Fig. 4 is a fragmentary longitudinal sectional view of a transmission embodying the invention showing the parts in a different position.

Fig. 5 is a fragmentary longitudinal sectional view of the prior art.

Fig. 6 is a fragmentary sectional view taken along the line 6-6 in Fig. 5.

Fig. 7 is a fragmentary longitudinal sectional view of a further form of the prior art.

Fig. 8 is a graph of tilt of rollers versus speed.

Fig. 9 is a fragmentary sectional view taken along the line 9-9 in Fig. 1.

Fig. 10 is a part sectional view of a portion of the structure shown in Figs. 1 to 9.

Figs. 11 and 12 are diagrams of the geometric development of the race contours of the transmission embodying the invention.

Fig. 13 is a longitudinal sectional view of a friction transmission embodying the invention, similar to Fig. 1 and showing additional features.

Fig. 14 is a sectional view taken along the line 14-14 in Fig. 13.

Fig. 1 5 is a sectional view taken along the line 15-15inFig. 13.

Fig. 1 is a cross section in a plane including the axis of the main shaft 106, a driven shaft 11 6 and a roller 103. A disc 101 is splined to shaft 106 and adjusted by nut 107. A disc 102 is mounted on shaft 106 by thrust bearing 143. Disc 101 has an annular race 31 of toric form facing a race 32 of like form in disc 1 02. A seat of rollers 103 contacts races 31 and 32 at contact points 30.

Each roller is held in a roller carrier 104 and rotates on bearing 105. As shown in Figs. 1 and 9, each roller carrier has trunnions 135 which are supported, as presently described with reference to Fig. 9, to permit the carrier with its roller to tilt as desired about an axis 37. This axis lies in a plane normal to the drive shaft axis 106 and is normal to the plane of the drawing in Fig. 1. Fig. 2 shows a roller 103 tilted about this axis.

As the driving disc 101 rotates in one direction in contact with the rollers 103, the driven disc 102 will rotate in the opposite direction at a speed depending on the tilt of the rollers 103. Means for controlling and changing the tilt are shown in Fig.

9.

An essential feature of the invention is that lines tangent to the contact surfaces of the rollers 103 and the races 31 and 32 converge at the intersection of the axes of rotation of the rollers 103 and shaft 106 at point 40. This resuits in true rolling action in the manner of cones with a common apex. This is true whatever the tilt angle as shown also in Figs. 2, 4 and 12.

Figs. 5 and 6 shows a prior art construction such as in United States Patent No. 1,844,464 and show the effect of other than true rolling. In such designs, the areas of contact of the roller 3a are directly opposite on a line through the tilt axis 37a. This results in an effect called "spin". It must be understood that there is no such thing as "point" contact. Any contact between two solid objects deforms the surfaces resulting in an "area" of contact. The size and shape of the deformed area depend upon the contours of the surfaces, the strength and elasticity of the materials and on pressure applied. Fig. 6 shows a circular contact 30a at a nominal distance 39c from the axis of shaft 6. This distance is presumed to establish the ratio between the race 31 a and the roller 3a.However, other points on this area are at different distances, as 39a and 39b, tending to establish different ratios. Therefore, there must be slippage at all points not lying on the arc of a single radius. Thus there is "spin" action, somewhat as if the race 31 a were turning on the roller around the center of the contact area.

This spin action has two serious effects: (1) Wear, unless good lubrication is provided, reducing the tractive force on which the transmission depends, and (2) induced slippage on much of the area, added to the tractive force reduces the effective tractive force and tends to cause gross slippage.

Fig. 7 shows a prior art form with a roller 3b with a convex conical roller as described in United States Patent No. 2,619,841, operating in races whose contours are circular arcs. Lines tangent to the contacting surfaces converge at points 40b which lie on a circular arc which crosses the axis of shaft 6 so that points 40b do not coincide with points 36b, the intersection of the axes. This reduces the spin but does not eliminate it. The angles 38a and 38b in Figs. 5 and 7 are measures of the spin.

A method of producing the race contours for true rolling action is disclosed in United States Patent No. 2,734,389. The present invention includes a combination of features necessary for the practical application of this idea.

In accordance with the invention as shown in Fig. 4, a roller with a convex conical rim operates in races 31 c and 32c whose contours are formed so that lines tangent to the contacting surfaces always converge at the intersection of the axes of rotation of roller 3c and shaft 6 as at points 36c and 40c.

Figs. 11 and 12 show the geometry of the roller 3 and the toric races 31 and 32, the axes of the drive shaft 6, the tilt axis 37 of the roller, the centers of contact 30 of roller and race, the intersection 36 of the roller axis and axis 6, and the converging point 40 on the roller axis of lines tangent to the contacting surfaces. In these diagrams points 36 and 40 coincide. These tangent lines represent elements of conjugate cones.

In the diagram, capital letters represent lengths of lines and small letters represent angles.

distance from drive shaft axis 6 to tilt axis 37; C=radius of the convex rim of the roller rim; a =angle from roller axis to a line through the tilt axis 37 and the center of radius C: B=distance from roller tilt axis to center of radius C.

The above values are constants. The following values vary with the tilt angle of the roller axis.

b tilt angle of the roller axis; E= distance on roller axis between points 37 and 40; E= A secant b; F= distance from point 40 to center of radius C; H and J are sides of a right triangle with angle a at point 37; J ""angle whose tangent = H E-J F=(E-J) secantj; c angle whose sine = C F c + j = angle between golfer axis and element of conjugate cone which is tangent to the contact surfaces; e =angle between the line tangent to the contact surfaces and a line parallel to the drive shaft axis 6 through point 37 or the angle from axis 6 itself; e =90 +(b-c-j); k =angle between a line through point 37 and the center of radius C, and a line also passing through the center of radius C perpendicular to the cone element at the point of tangency; k=900 -(a + c + j).

Angle k represents the shift of center of contact 30 on the roller rim as the roller is tilted.

If this line perpendicular to the cone element at the point of tangency is extended through the center of radius C to intersect the roller axis at points P, then point P is the instantaneous center of tilt instead of the mechanical axis at point 37.

Point P shifts as the roller is tilted, moving above or below axis 37 as angle k changes, becoming positive or negative.

Points on the toric curve may be located as junctions of successive chords. The angle ec of a chord is the mid-value between angles en and en+ 1 derived from small increments of angle b.

Thus

As shown in Fig. 11, XO and YO are the values at b =o.

Xo=B sine a + C sine ego).

Y0 = -(B cosine a + C cosine ego).

Fig. 12 shows the roller tilted from the central position by an increment A b. In this case en is eO.

The length K of each chord is K=2B sine Ab 2 cosine kc + 2C sine Ae 2 as shown in the diagram where L=the first term and 2M the second term. Here the angle kc is the value related to ec.

Increment A X = K cosine ec and A Y = K sine ec.

Then

Thus the curve starts at the zero or mid-position of the roller, and is generated by successive chords at angle e following the contour of the roller rim as the roller tilts in either direction approaching the desired curve as a limit. This accomplishes the primary objective of aligning the contacting surfaces with elements of cones having a common apex point.

The X-Y coordinates of the curve may be computed with any desired precision by solving the above equations based on small increments of angle b.

Referring again to Fig. 1, shafts 106 and 11 6 are supported in bearings 111 and 113 which are mounted in housing 1 66. Nut 107 adjusts discs 101 and 102 to a fixed and precise spacing.

Bearing 111 is designed to keep the discs 101 and 1 O2 in proper relation to the rollers 103 as thrust forces vary with the tilt position of the rollers. Pressure on the contact surfaces 30 produces a thrust force along the axis of the roller 103 which bearing 105 is designed to carry as well as the radial load produced by the tractive forces. Bearing 143 is designed to carry the forces tending to separate discs 101 and 102 in the direction of the drive shaft axis 106.

As previously noted, the tilt axis is displaced slightly if the instantaneous tilt center P does not coincide with the mechanical tilt axis 37 as the roller 103 is tilted. Means as shown in Fig. 9 must be provided in the mechanism for controlling and positioning the rollers 103 to permit this displacement and at the same time maintain adequate pressure in the direction of the roller axis' to create the necessary traction. This pressure must be applied in the direction of the roller axis at an angle in order to equalize the contact pressures at points 30 on the roller. This cannot be accomplished by applying pressure in the direction of the drive shaft 106 because of the difference in the contact angles on discs 101 and 102.At extreme angles of tilt, the contact angle measured from the axis of shaft 106 may be as low as 1 50 while the other is 55 . The contact pressure is inversely proportional to the sine of the angle, so that the pressure at the smaller angle might be three times as great as the other.

Fig. 1 shows a bearing 105 for roller 103 in which the inner race 148 may move in a direction normal to the roller axis in a plane radial to the main shaft axis 106. A pressure spool 108 has a tongue 1 08A which projects into a groove in inner race 148. A plug 1 69 closes the end of this groove. The pressure spool 108 is aligned by multiple diaphragm springs 149 and 1 52. Such springs can be designed to apply nearly uniform load within a range which will accommodate the axial component of the displacement of the tilt center, In some applications where loading is nearly uniform, the use of such springs alone may be satisfactory.However, for applications such as vehicle propulsion where there is extreme variation in loading, if the pressure were maintained at the maximum at all times, the durability of bearings and other parts would be reduced.

Figs. 1 and 9 show a design where the major part of the load on the rollers 103 is supplied by hydraulic pressure in proportion to the applied driving load. The pressure spool 108 is the core of a piston. A small initial load is applied by the diaphragm springs. In Figs. 1 and 9, springs 1 52 are sealed by a rubber seal 1 53 and held by retaining ring 1 54. The springs add to the effective area of the piston. The piston is held in axial alignment by the closely fitted springs. Two guide pins 1 55 prevent rotation. The roller carrier trunnions 135 are mounted in bearings consisting of rollers 156 and pistons 128 and 129 which, in turn, are supported and centered by diaphragm springs 1 50. Balls 1 59 form thrust bearings.

Springs 1 50 are held in bracket 127 by retaining ring 125 and sealed by rubber seal 151. An extension on each piston fits into a bore in bracket 127 to maintain alignment and to act as a dashpot. A small orifice 76 reguiates the action of the dashpot. Seal 157 and spacer 1 58 confine rollers 1 56.

The pistons move the roller carrier in a plane normal to main shaft axis 106 so that the roller axis is bffset by a controlled amount 67 from the main shaft axis 106 as shown in Fig. 10. The bearing rollers 1 56 permit the roller carrier to tilt and the diaphragm springs 1 50 permit it to shift axially with a minimum of static friction.

In Fig. 10, if disc 101 is considered to rotate in a clockwise direction and the roller carrier with its roller 103 is shifted to the left, the roller in contact at point 30 will follow a spiral path, increasing the radial distance 68 from the center of disc 101.

Conversely, the radial distance of the opposite contact of the roller on disc 102 will be decreased.

This spiral is an involute with a base radius equal to the offset 67 so that the advanve for each revolution is 2 or times the offset. Obviously, this shifting may be quite rapid.

If the speed of revoiution of disc 101 is substantially constant, as with a flywheel and the output shaft is connected, for example, with the drive wheels of a vehicle, the driving force reacting on the roller 103 will increase, tending to move the roller carrier 104 to the right, reducing or stopping the tilting action. Hydraulic pressure on the piston 128 is balanced by the reactive force on the roller 103, thus providing a controllable means of increasing or decreasing the ratio. In the same manner, pressure on the opposite piston 129 may be used to balance the reactive force on roller 103 to decelerate the vehicle or control its speed on a downgrade. Fig. 9 shows three rollers 103 and three pairs of pistons. Ports 141 transmit pressure to the "driving" pistons and ports 142 to the "retard" pistons, each set of ports being connected by a manifold to control valve 170.

In Fig. 9, pressure from a pair of control pistons 128 and 129 is transmitted through connector tubes 1 62 to the roller pressure piston 108 proportional to the driving load. A shuttle valve 1 60 admits fluid from the high pressure side but closes the other side bv contact with aPat 1 S1 on that side. Each connector tube is sealed by ring seals 1 63 in the piston and in the roller carrier.

It is especially important to eliminate static friction in the axial shifting of the roller carrier in order to obtain smooth and accurate ratio control.

As shown in Fig. 9, a pump 171 supplies hydraulic fluid under pressure to a valve assembly 1 70 which controls the pressure delivered to manifold 141A which is connected to "drive" ports 141 and to manifold 1 42A connected to "retard" ports 142. The pump 1 71 draws fluid from reservoir 174 through tube 172. Low pressure fluid is returned to the reservoir through tube 173.

In a vehicle propulsion system, for example, the valve asembly 1 70 may receive signals through a number of connections 175, such as vehicle driver control, engine speed, and engine torque as determined by reaction pressure from manifolds 141A and 142A in connection with the transmission ratio as determined by the roller tilt position. This invormation may be organized according to a program which will control the speed and torque of the engine for maximum fuel economy. In other applications, other data might be important for control.

Again in Fig. 1, a static seal 109 contacts disc 101 and shaft 106, a running seal 123 mounted in disc 102 contacts shaft 106, a seal 120 is mounted in housing 1 66 to seal bearing spacer 110 and a seal 124 in roller 103 seals inner race 148. The rim of disc 102 is shaped to act as a baffle and slinger in conjunction with the shape of housing 166.

Although the space in which the friction members operate is sealed and drained, some contamination of the friction surfaces may occur over a long period of time in the absence of an oil bath. Accordingly, this invention provides a means of removing this contamination.

Also, it may be possible to increase the coefficient of friction by applying coating materials on the friction surfaces, perhaps in combination with cleaning materials.

If slipping at the traction contacts should occur without lubrication, it might be very destructive.

This might be caused by transient impulses of high torque in the drive line of a vehicle due to road irregularities. A bump compresses the tire of a drive wheel, reducing its radius and accelerating its rotation. This acceleration is resisted by the inertia of the engine. A similar effect might be produced by the rhythmic bouncing of the wheel.

There are doubtless other transient and abnormal conditions which might cause impulses of excess torque.

As described previously with reference to Figs.

1 and 9, the pressure on the roller piston is controlled by the drive torque through hydraulic pressure in one form of the invention. Thus, the pressure on the traction contacts would be low at low drive loads. Hydraulic pressure would increase in response to an increase in torque, however, it might not be quick enough to prevent momentary slipping. The diaphragm springs which align the roller pistons provide an initial load on the rollers which adds a factor of safety against slipping at low drive loads.

In another form in the invention, as shown in Figs. 13 and 14, pressure on the rollers is produced by springs only. In this form, the effective radius on the driven torus might be small due to a high ratio (overdrive) position of the rollers. The transmission would be designed to deliver maximum drive wheel torque only at a low ratio where the effective radius on the driven torus would be near maximum so that the roller pressure might not be sufficient to prevent slipping at the small effective radius on the driven torus.

Accordingly, it is necessary to provide a torque limiting clutch in the drive line. The slipping torque of the clutch must be moduiated to transmit full engine torque at any ratio, but slip when the torque being transmitted at any moment is suddenly increased.

Valve assembly 140 in Fig. 14 performs another function in addition to ratio control, that of controlling hydraulic pressure on the protective clutch in the drive line. Pipe 100 is connected to pipe 10Q in Fig. 13 and delivers pressure proportional to the torque on the shaft in which the clutch is installed.

The protective clutch is shown in Fig. 13 where output shaft 126 is splined to a clutch casing 96 which contains a piston 97. The driving plates of a multiple disc clutch 98 are splined to casing 96.

The driven plates of clutch 98 are splined to a secondary output shaft 99. Fluid pressure is delivered by pipe 100 and sleeve 95 to piston 97.

Seal rings 121,122 and 132 prevent loss of pressure.

While the drawing shows the protective clutch on the output shaft of the transmission, it can be on the input shaft. The multiple disc clutch might be replaced by any other device which would permit relative rotation or "slip" when subjected to torque in excess of a controlled amount. The term "protective clutch" is intended to include any such device.

The slipping torque of the protective clutch must be greater than the torque delivered by the engine or transmission but less than that which would cause slipping at the traction contacts of the transmission. Because there will be considerable variation in the coefficient of friction of the plates of the multiple disc clutch, a factor of safety must be allowed in the design coefficient of friction of the traction contacts of the transmission.

Fig, 1 3 shows a cleaning and coating system for the friction surfaces comprising a pump 176 delivering a cleaning fluid or coating material through pipes 1 78 and 179 to jets 1 80. Fluid or other material is drawn through pipe 1 77 from reservoir 1 88 and drained from case 1 65 through pipe 181. Pump 182 draws vapors from case 165 and delivers them through pipe 183 to the atmosphere or to the emission control system.

Rotation of members 101 and 102 and tilting of rollers 103 with the clutch disengaged distributes the material over the friction surfaces.

Suitable cleaning agents might be solvents, solutions of detergents, suspensions of solid particles or mildly corrosive solutions. Suitable coating materials might be ultra thin films of resin, lacquer, metal powder or other.

The cleaning and coating operations may be done periodically on a time or mileage schedule with either manual or automatic control or a continuous process may be used.

Figs. 13 and 14 show a roller spindle 138 replacing the roller piston 108. Multiple diaphragm springs 1 84 provide a substantially uniform load on the roller. Spacers 1 85 and 1 86 separate two groups of springs 1 84 to provide a wide base for aligning spindle 138. A threaded hole in the spindle permits the use of a retracting screw during assembly. A plug 187 in carrier 164 closes an access hole. Carrier 1 64 and pistons 130 and 131 do not have the connecting holes for hydraulic control of roller pressure.

Fig. 1 5 is a cross section of the roller and bearing and shows extension 1 38A of the roller spindle in inner race 148 permitting lateral movement. Extension 1 08A in Fig. 1 is similar to extension 138A.

Disc 102 has an extension 144 with spline teeth 145. Slideably mounted on these spline teeth is a shifter collar 115 with engaging teeth 146. In a set of planetary gears, pinions 11 7 are mounted on shafts 11 8 secured in a cage on the end of shaft 11 6. The pinions are held in place by thrust washers 73. On the flange of shaft 116 are teeth 133 engageable by teeth 146 on shifter collar 11 5. Rotatably mounted in extension 1 44 of disc 102 is a sleeve 114 having a flange with teeth 1 34 engageable by teeth 146 of shifter collar 11 5. On the inner surface of sleeve 114 are gear teeth engaging the teeth of planet pinions 117. Sleeve 114 is held in place by thrust washers 74 and 75. On an extension of shaft 106 are gear teeth 147 engaging pinions 11 7. A further extension of shaft 106 fits into a bearing 119 contained in shaft 116. Fig. 3 shows a transverse cross section of these planetary gears.

When the shifter collar 11 5 is in the position shown, sleeve 114 must rotate with disc 102, operating the planetary gears. This reduces the speed of output shaft 11 6 in relation to the speed of disc 102 as driven by rollers 103 and disc 101.

When the ratio between disc 102 and disc 101 is reduced to the ratio of the planetary gears by tilting the rollers 103 toward the position shown in Fig. 2, the rotation of shaft 11 6 is stopped. This provides a neutral or stationary relation. If the rollers are tilted further to reduce the ratio between disc 102 and disc 101, the rotation of shaft 11 6 is reversed. If the rollers are tilted from the neutral position to increase the ratio between disc 102 and disc 1 01, a forward motion of shaft 11 6 is obtained. Thus, a vehicle may be started, stopped or reversed smoothly under full control of the transmission.

As the forward motion is increased, a speed is reached at which the shifter collar 11 5 may be moved to the right to engage teeth 133 on the flange of shaft 11 6, connecting it directly to disc 102 and idling the planetary gears. When this shift is made, it is necessary to tilt the rollers to obtain a suitable ratio between disc 102 and disc 101.

Fig. 8 shows a possible example. In the diagram, the tilt angle of rollers 103 is shown as scale 22. A hypothetical scale of miles per hour is shown as 70. Curve 71 is the speed versus tilt relation with the planetary gears operating. Curve 71 A is the reverse portion. Curve 72 is the speed versus tilt relation with shaft 11 6 connected directly to disc 102. 26 indicates the range of overlap of the speed curves in which a shift may be made without changing the speeds of the driving or driven devices.

A simple sliding collar type of shifter is shown for illustration. Synchronizing devices or friction clutches in highly developed form may be used.

This invention provides a continuously variable friction transmission with true rolling action which does not require lubrication of the traction surfaces. This greatly reduces the required contact pressures and permits the use of conventional thrust bearings of reasonable size. It is adaptable to any degree of manual or automatic control.

Claims (14)

1. In a friction transmission, the combination comprising a driving member, a driven member, a common shaft, means of supporting said members for rotation about the axis of said common shaft, said members having oppositely facing annular arcuate surfaces, a thrust bearing on said common shaft, one of said members being mounted on said thrust bearing on said common shaft, means for fixing said members axially so that there is an exact spacing between said annular surfaces, at least one roller contacting said surfaces, means for supporting said roller with the axis thereof intersecting the common axis of said members, said means of supporting said roller having a tilt axis in a plane normal to the common axis of said members, the periphery of said roller having a generally convex conical configuration, said annular surfaces of said members having a configuration such that lines tangent to the periphery of the roller and the annular surfaces converge at the intersection of the axis of rotation of the roller and the common axis of the members at all positions of the roller.

said means of supporting said roller permitting axial and lateral movement to conform to the noncircular contour of the annular surfaces, means of applying a force on said roller in the direction of the roller axis, and means for maintaining said tilt axis in a fixed iocation with respect to the axial location of said members resulting in equal contact pressures of the roller on both members.

2. The combination set forth in claim 1 wherein said means for supporting said roller includes trunnion means for supporting said roller for rotation about an axis transverse to the axis of said roller and piston means associated with said trunnion means urging said trunnion means in the direction of said transverse axis.

3. The combination set forth in claim 1 wherein said means for supporting said roller comprises hydraulic means for applying a force proportional to the driving load on said piston means to urge said roller in the direction of the roller axis, and diaphragm means supporting said trunnion means for movement along said transverse axis only.

4. The combination set forth in any of claims 1 or 2 including means of applying a fluent material on the surfaces.

5. The transmission set forth in claim 4 wherein said material comprises a material to maintain or increase the coefficient of friction.

6. The transmission set forth in claim 5 including means for removing said material from the friction surfaces.

7. The transmission set forth in any of claims 1-6 including a protective clutch coupled to said transmission to prevent slipping of the friction members.

8. The transmission set forth in claim 7 wherein the slipping torque of said protective clutch is controlled in relation to the normal transmitted torque, said protective clutch to slip when the transmitted torque is increased by transient or abnormal conditions.

9. The transmission set forth in any of claims 1-8 including means for removing said fluent material.

1 0. The transmission set forth in claim 9 wherein said means for applying and means for removing said material comprises pump means.

11. In a variable ratio friction transmission comprising coaxial driving and driven members with opposed toric races having friction surfaces and interposed convex conical rollers, the races and rollers being of such form as to provide true rolling contact in the manner of cones with a common apex and with substantially dry frictional contact, the method comprising applying a fluent material to said friction surfaces of said transmission after the transmission is assembled.

12. The method set forth in claim 11 including the step of removing said fluent material from said surfaces after the transmission is assembled.

13. A friction transmission constructed and arranged substantially as hereinbefore described with reference to and as illustrated in the accompanying drawings.

14. A method of constructing a friction transmission substantially as hereinbefore described with reference to the accompanying drawngs.