Power roller for a toroidal continuously variable transmission

Abstract: A power roller for a toroidal continuously variable transmission is provided wherein decrease in power transmission performance due to oil film temperature rise in a power transmission part can be prevented by suppressing the heat in a rolling contact part between a ball and an inner race groove. In a thrust ball bearing of the power roller, by setting a radius of curvature r i of an inner race groove (67c) less than a radius of curvature r o of an outer race groove (67d), a Hertzian contact area between the inner race groove (67c) and a ball (67a) is set larger than a Hertzian contact area between the outer race groove (67d) and the ball (67a). By so doing, the direction of the rotation axis of the ball (67a) can be set to a direction wherein the spin of the inner race is less than the spin of the outer race.

Description: This invention relates to a power roller for a toroidal continuously variable transmission according to the preamble portion of claim 1. A power roller according to the preamble portion of claim 1 is known from US 5,536,091. In recent years, research has been carried out on types of vehicle transmission which continuously vary a speed ratio. In addition to belt type continuously variable transmissions, traction drive continuously variable transmissions which transmit power by the shear of oil film (referred to hereafter as toroidal CVT) are well known. Toroidal CVT allow an engine to be driven at optimum efficiency, and this technique is therefore of much interest as it permits a large improvement of fuel cost performance and power transmission. Jikkai Sho 62-199557 published by the Japanese Patent Office in 1987 discloses such a toroidal CVT. According to Jikkai Sho 62-199557, a prior art toroidal CVT comprises a power roller having an inner race which transmits power between an input disk and an output disk via an oil film, an outer race supported by a trunnion (power roller supporting member), and a thrust ball bearing which supports the inner race on the outer race so that is free to rotate. This thrust ball bearing comprises plural balls, a holder which holds these balls, and an inner race groove and outer race groove formed respectively in opposite surfaces of the inner race and outer race. Both of these race grooves have an arc-shaped cross-section, and the whole forms a ring. In the prior art, a radius of curvature r i of the inner race groove and a radius of curvature r o of the outer race groove are identical, and the pitch diameter d i of the inner race groove and the pitch diameter d o of the outer race groove are also identical. However, in the prior art toroidal continuously variable transmission, the direction of the rotation axis of the balls of the thrust ball bearing cannot be controlled. Therefore, if the vehicle is driven for a long-time while the spin angular velocity of the balls relative to the inner race groove is large, the rolling contact parts of the balls and inner race groove generate heat, and the temperature of the inner race groove rises. The inner race transmits power from the input disk to the output disk by traction drive. A traction coefficient degreases together with temperature rise of an oil film between the inner race and input disk or between the inner race and output disk, and when the temperature of the inner race rises, power transmission efficiency declines. It is therefore an object of this invention to provide a power roller as indicated above which can prevent decrease of power transmission efficiency due to temperature rise of an oil film of a power transmission part. This objective is solved by a power roller having the features of claim 1. In the following, the present invention is explained in greater detail with reference to the accompanying drawings, wherein: FIG. 1 is a schematic view of an embodiment of a toroidal continuously variable transmission, FIG. 2 is a schematic view showing a speed change control system used with the toroidal continuously variable transmission, FIG. 3 is a partial sectional view of a power roller wherein the inner race groove has a radius of curvature less than the radius of curvature of the outer race groove, not underfalling the scope of claim 1, FIG. 4 shows the relation between a rotation axis angle of a ball of a thrust ball bearing and an inner race spin moment, the relation between a rotation axis angle of a ball of a thrust ball bearing and an outer race spin moment, and the relation between the rotation axis angle of the ball and a power loss, for the power roller of FIG. 3. FIG. 5 is a partial sectional view of a power roller wherein the inner race groove has a pitch diameter less than the pitch diameter of the outer race groove, not underfalling the scope of claim 1, FIG. 6 shows the relation between a rotation axis angle of a ball of a thrust ball bearing and an inner race spin moment, the relation between a rotation axis angle of a ball of a thrust ball bearing and an outer race spin moment, and the relation between the rotation axis angle of the ball and a power loss, for the power roller of FIG. 5, FIG. 7 is a partial cross-sectional view showing a power roller according to the invention, FIG. 8 shows the relation between a rotation axis angle of a ball of a thrust ball bearing and an inner race spin moment, and the relation between a rotation axis angle of a ball of a thrust ball bearing and an outer race spin moment, in a power roller according to the invention of FIG. 7, and FIG. 9 shows the relation between a rotation axis angle of the ball and an inner race spin moment, the relation between a rotation axis angle of the ball and an outer race spin moment, and the relation between the rotation axis angle of the ball and power loss, according to the power roller of the prior art. Referring to FIG. 1 to Fig. 2 the mechanical construction of a half toroidal continuously variable transmission (referred to hereafter as toroidal CVT) having a dual cavity will first be described. The toroidal CVT comprises an input shaft 4 connected to an output shaft 2 of an engine 1 via a torsional damper 3, and a hollow shaft 5 formed coaxially on the outside of the shaft 4 and supported free to rotate relative to it. A first toroidal CVT 6 and second toroidal CVT 7, and a loading cam 8 which presses input disks 61, 71 with a load which is directly proportional to an input torque, are disposed coaxially on these shafts 4,5. The first toroidal CVT 6 and second toroidal CVT 7 have a substantially identical construction. These elements comprise the two input disks 61, 71 and two output disks 62, 72 whereof the opposite surfaces form a toroidal space with an arc-shaped cross-section. Two power rollers 63 which transmit power are held between the input disk 61 and output disk 62. Two power rollers 73 which transmit power are held between the input disk 71 and output disk 72. The first toroidal CVT 6 arranged on the far side of the engine comprises the input disk 61 on the opposite side to the engine, and the output disk 62 on the engine side. The second toroidal CVT 7 disposed near the engine 1 comprises the input disk 71 on the engine side and the output disk 72 on the opposite side to the engine. The two input disks 61, 71 are respectively slidably joined to the two ends of the hollow shaft 5. The two output disks 62, 72 (referred to as one-piece output disk 60) are formed in one piece, and supported free to rotate in the middle part of the hollow shaft 5. One of the power rollers 63 comprises an inner race 64 which transmits power from the input disk 61 to the output disk 62 via an oil film, an outer race 66 supported by a trunnion 65 (power roller supporting member), and a thrust ball bearing 67 which supports the inner race free to rotate relative to the outer race 66. One of the power rollers 73 comprises an inner race 74 which transmits power from the input disk 71 to the output disk 72 via an oil film, an outer race 76 supported by a trunnion 75 (power roller supporting member), and a thrust ball bearing 77 which supports the inner race free to rotate relative to the outer race 76. An output gear 9 is provided on the outer circumference of the one-piece output disk 60. The output gear 9 meshes with a drive gear 10 of a differential device 11. This transmits drive force to left and right drive wheels via drive shafts 12a, 12b extending to the left and right from the differential device 11. The oil pressure system for managing the speed change of the first toroidal CVT 6 and the speed change system of the first toroidal CVT 6 will now be described referring to FIG. 2. Speed change of the first toroidal CVT 6 is performed by moving the trunnion 65 which supports the power roller 63 up and down from a balance point. Due to this displacement, at the contact point, a difference arises between the vectors of the rotation directions of the inner race 64 and input disk 61 and between the vectors of the rotation directions of the inner race 64 and output disk 62, so the power roller 63 gyrates. The trunnion 65 is connected to a servo piston 31 of an oil pressure servo 30, and displaces due to the differential pressure between the oil in a cylinder 30a on the high side and the oil in a cylinder 30b on the low side. The oil pressure of the high side cylinder 30a and the oil pressure of the low side cylinder 30b are controlled by a shift control valve 46. Due to the displacement of a spool 46S in the valve, the shift control valve 46 allows oil supplied from a line pressure port 46L to flow to a high side port 46Hi or a low side port 46Low. Oil then flows from the other of the low side port 46Low or high side port 46Hi to a drain port 46D. Thus, the differential pressure in the oil pressure servo 30 varies. A precess cam 35 is fitted to one of the trunnions 65, and a groove is cut in the precess cam 35. The groove in the precess cam 35 is in contact with one end of an L link 38, and the end of the L link 38 is freely supported by one end of an I link 37. Hence, the displacement and gyration angle of the trunnion 45 is fed back to the I link 37. The other end of the I link 37 is connected to a step motor 36, and the spool 46S of the shift control valve 46 mentioned above is freely supported on the I link 37. Therefore, the displacement of the spool 46S is determined from the displacement of the step motor 36 and feedback from the precess cam 35. Referring to FIG. 3 and FIG. 4, the effect of making the radius of curvature r i of the inner race groove 67c smaller than the radius of curvature r o of the outer race groove 67d in the power roller 63 of the toroidal CVT according to the first embodiment will now be described. Here the pitch diameter d i of the inner race groove 67c is set to be equal to the pitch diameter d o of the outer race groove. In FIG. 3, reference numeral 63 denotes the power roller, reference numeral 64 denotes the inner race, reference numeral 66 denotes the outer race and reference numeral 67 denotes the thrust ball bearing. The thrust ball bearing 67 comprises plural balls 67a, a holder 67b which holds these balls 67a, an inner race groove 67c formed on the surface of the inner race 64 facing the outer race 66, and an outer race groove 67d formed on the surface of the outer race 66 facing the inner race 64. The inner race groove 67c and outer race groove 67d have arc-shaped cross-sections, and define an annular space. The plural balls 67a are situated in the annular space, and come in contact with the inner race groove 67c and outer race groove 67d. Each of the plural balls 67a revolves around the rotation axis 100 of the power roller 63, rotating about a rotation axis 50, because of the rotation of the inner race 64. The holder 67b revolves around the rotation axis 100 of the power roller 63, together with the balls 67a. Referring to FIG. 3, a rotation axis angle γ is defined as the angle formed by the rotation axis 50 of the ball 67a and a plane perpendicular to the rotation axis 100 of the power roller 63. In other words, a rotation axis angle γ is defined as the angle between a rotation axis 50 of the ball 67a and the tangential plane on the inner race side of the ball 67a perpendicular to a rotation axis 100 of the power roller 63, as shown in FIG. 3. In FIG. 3, when the rotation axis 50 of the ball 67a inclines to the lower right, the rotation axis angle γ takes a positive value. In other words, when the rotation axis 50 of the ball 67a inclines towards the rotation axis 100 of the power roller 63 and toward the inner race 64, the rotation axis angle γ takes a positive value. Describing this in still other terms, the rotation axis angle γ takes a positive value when the intersection point between the rotation axis 100 of the power roller 63 and the rotation axis 50 of the ball 67a lies on the side of the inner race 64 with respect to the revolution plane 150 of the balls which passes through the centers of the balls 67a and is perpendicular to the rotation axis 100 of the power roller 63. When the rotation axis 50 of the ball 67a is perpendicular to the rotation axis 100 of the power roller 63, the rotation axis angle γ is zero. In FIG. 3, when there is an inclination to the upper right, the rotation axis angle γ takes a negative value. In other words, when the rotation axis 50 of the ball 67a extends simultaneously towards the rotation axis 100 of the power roller 63 and the outer race 66, the rotation axis angle γ takes a negative value. The ball 67a rotates as it slides on the inner race groove 67c. In the rolling contact part between the ball 67a and the inner race groove 67c, sliding between the ball 67a and inner race groove 67c in such a manner that the ball 67a and inner race groove 67c rotate relative to one another is defined as an inner race spin (first spin). The frictional moment associated with the inner race spin is referred to as an inner race spin moment (first spin moment) M si . The ball 67a also rotates as it slides on the outer race groove 67d. In the rolling contact part between the ball 67a and outer race groove 67d, sliding between the ball 67a and outer race groove 67d in such a manner that the ball 67a and outer race groove 67d rotate relative to one another is defined as an outer race spin (second spin). The frictional moment associated with the outer race spin is referred to as an outer race spin moment (second spin moment) M so . The inner race and outer race spins cause power loss and result in the generation of heat. As described on page 119 of "Roller Bearing Engineering" (published by Yokendo, in Tokyo, in 1978), the spin angular velocity on the side of the inner race 64 and outer race 66 depends on the rotation axis angle γ of the ball 67a. According to this reference, an inner race spin angular velocity ω si and outer race spin angular velocity ω so are defined by the following equations: ω si =(ω i -ω c )sinα+ω b sin(-γ) ω so =ω c sinα+ω b sinγ Here, ω i is the angular velocity of the inner race, ω c is the revolution angular velocity of the ball, ω b is the rotation angular velocity on the ball, α is the contact angle (this is the angle subtended by a straight line joining the contact point of the ball and inner race and the contact point of the ball and outer race, with the revolution plane 150). Normally, ( ω i -ω c ) is equal to ω c . If the rotation axis angle γ is positive, the inner race spin angular velocity ω si is less than the outer race spin angular velocity ω so . If the rotation axis angle γ is negative, the inner race spin angular velocity ω si is larger than the outer race spin angular velocity ω so . According to page 696 of "Transactions of the ASME Vol. 1100: October 1988", the rotation axis angle γ is effectively determined by the balance between the first spin moment M si and second spin moment M so . Specifically, the rotation axis angle γ is determined under the condition M si ≅M so . By making the radius of curvature r i of the inner race groove 67c smaller than the radius of curvature r o of the outer race groove 67d, the Hertzian contact area between the inner race groove 67c and ball 67a is set larger than the Hertzian contact area between the outer race groove 67d and ball 67a. Thus, due to a reason described later, the rotation axis angle γ of the ball 67a can be set to a rotation axis angle γ2 which is a positive value. Specifically, the rotation axis 50 of the ball 67a is set in a direction pointing towards-both the rotation axis of the power roller 63 and the inner race 64. At this rotation axis angle γ2 , the inner race spin angular velocity ω si is smaller than the outer race spin angular velocity ω so . Consequently, the heat generated in direct proportion to the product of the spin moment and spin angular velocity is suppressed smaller on the inner race side. Also, decrease of the traction coefficient accompanying increase of oil film temperature is suppressed, and decrease of the power transmission performance due to the ball spin of the thrust ball bearings 67, 77 can be prevented. In the power roller 73 also, the rotation axis angle γ of the ball 67a can be set to the rotation axis angle γ2 at which the inner race spin angular velocity ω si is less than the outer race spin angular velocity ω so , in the same way as for the power roller 63. The reason why the rotation axis angle γ can be set to the rotation axis angle γ2 at which the inner race spin is less than the outer race spin, will now be described in detail referring to FIG. 4. FIG. 4 shows the characteristics of the inner race spin moment M si , the inner race spin moment M si of a prior art example (where r i =r o ), the outer race spin moment M so and the power loss characteristics. These parameters were calculated using an elastic model according to "Proceedings of the Mechanical Engineering Institute of Japan", No. 89-0148B (1989), 512B. The spin moment is obtained from a spin angular velocity depending on the rotation axis angle γ. The rotation axis angle γ of the ball 67a is determined by the balance between the inner race spin moment M si and outer race spin moment M so . As shown in FIG. 4, at a first rotation axis angle γ0 , second rotation axis angle γ1 and third rotation axis angle γ2 , the spin moments (frictional moments) M si , M so of the inner and outer races are balanced. The power loss is given by M si •ωsi+ M so •ωso. At the third rotation axis angle γ2 where the inner race spin angular velocity ω si is less than the outer race spin angular velocity ω so , the power loss is smallest. Therefore, during running of the CVT, the third rotation axis angle γ2 at which the power loss is minimized, is maintained. For example, even if the rotation axis angle γ instantaneously becomes the first or second rotation axis angles γ0 , γ1 , the rotation axis angle γ immediately returns to the third rotation axis angle γ2 at which the power loss is a minimum. In the prior art toroidal CVT, unlike the case of this invention, the radius of curvature r i of the inner race groove 67c and the radius of curvature r o of the outer race groove 67d are set equal. Consequently, as shown in FIG. 9, in the prior art toroidal CVT, the absolute values of the second rotation axis angle γ1 and the third rotation axis angle γ2 are equal. In the prior art toroidal CVT, as shown in FIG. 9, the rotation axis angle γ at which the spin moments M si , M so are equal are the three points γ1, γ0 and γ2. However, at the point where γ=γ0 , the power loss is larger than at the other two points, so γ= γ0 cannot be maintained. Therefore, the rotation axis angle γ takes a value of either γ1 or γ2 with a probability of 50%, and the rotation axis angle γ cannot be controlled. During running of the toroidal CVT, the rotation axis angle γ varies between γ1 and γ2 . On the other hand, in this invention, by making the radius of curvature r i of the inner race groove 67c smaller than the radius of curvature r o of the outer race groove 67d, due to a reason described later, the Hertzian contact area between the inner race groove 67c and the ball 67a can be set larger than the Hertzian contact area between the outer race groove 67d and the ball 67a. The spin moment is directly proportional to the magnitude (moment span) of the Hertzian contact area. Therefore, the spin moment M si on the inner race side when the rotation axis angle is γ , is larger than the spin moment M so when the rotation axis angle is -γ. As shown in FIG. 4, at a rotation axis angle γ2 which is larger than in the prior art (i.e., r i =r o ), the inner spin moment M si and outer spin moment M so balance. Thus, as the power loss is minimized, the direction of the rotation axis of the ball is stabilized at the rotation axis angle γ2 . By setting the Hertzian contact area between the inner race groove 67c and ball 67a, to be larger than the Hertzian contact area between the outer race groove 67d and ball 67a in this way, the direction of the rotation axis is stabilized at the positive rotation axis angle γ2 . Here, the reason that the Hertzian contact area of the inner race is larger than the Hertzian contact area of the outer race, will be described. The principal curvature ρ (inverse 1/ R of radius of curvature R) in the contact part between the inner race 64 and ball 67a, and contact part between the outer race 66 and ball 67a, is given by the following equations. ρ 11i =ρ 12i =ρ 11o =ρ 12o =2/Da ρ 21i =-1/r i ρ 21o =-1/r o ρ 22i =ρ 22o =0 Here, Da is the diameter of the ball 67a. ρ 11i is the primary principal curvature of the ball 67a at the contact point on the inner race side, ρ 12i is the secondary principal curvature of the ball 67a and contact point on the inner race side, ρ 11o is the primary principal curvature of the ball 67a at the contact point on the outer race side, ρ 12o is the secondary principal curvature of the ball 67a at the contact point on the outer race side, ρ 21i is the primary principal curvature of the inner race groove 67c, ρ 21o is the primary principal curvature of the outer race groove 67d, ρ 22i is the secondary principal curvature of the inner race groove, and ρ 22o is the secondary principal curvature of the outer race groove. To compare the Hertzian contact area of the inner race 64 with the Hertzian contact area of the outer race 66, the values of ρ 21i and ρ 21o are compared. For simplicity, considering the case where the cylindrical inner surfaces of curvature ρ 21i and curvature ρ 21o come in contact with a cylinder of radius of curvature R , the respective combined curvatures Σρ i , Σρ o are given as follows: Σρ i =(1/R)-(1/r i ) Σρ o =(1/R)-(1/r o ) Here, the combined curvature Σρ is the curvature when contact between cylinders is replaced by contact between a cylinder and a plane. From the above equations, Σρ i <Σρ o , so the cylinder having a curvature of ρ 21i has a larger contact area than the cylinder having a curvature of ρ 21o . Therefore, also in the contact between the groove and the ball, the inner race has a larger contact area. Next, the effect of making the pitch diameter d i of the inner race groove 67c smaller than the pitch diameter d o of the outer race groove 67d will be described referring to FIG. 5 and FIG. 6. Referring to FIG. 5, in the thrust ball bearing 67, the inner race groove 67c and outer race groove 67d have an identical radius of curvature r. However, the pitch diameter d i of the inner race groove 67c is smaller than the pitch diameter d o of the outer race groove 67d. Here, the pitch diameter  d i of the inner race groove 67c is twice the distance from the center of the arc-shaped cross-section of the inner race groove 67c to the rotation axis of the power roller 63. The pitch diameter d o of the outer race groove 67d is twice the distance from the center of the arc-shaped cross-section of the outer race groove 67d to the rotation axis of the power roller 63. Hence, the Hertzian contact area between the inner race groove 67c and the ball 67a, is set larger than the Hertzian contact area between the outer race groove 67d and the ball 67a. Due to the relation between the pitch diameter d i of inner race and the pitch diameter d o of the outer race, the contact angle α is larger than 90 degrees when viewed from the inner race 64. At this time, the principal curvature in the contact part between the inner race 64 and ball 67a, and in the contact part between the outer race 66 and ball 67a, is given by the following equations. ρ 11i =ρ 12 i =ρ 11o =ρ 12o =2/Da ρ 21i =ρ 21o =1/r ρ 22i =2cosα/(d i -Dacosα) ρ 22o =-2cosα/(d o +Dacosα) Here, ρ is the curvature of the groove, α is the contact angle, Da is the ball diameter. As α is larger than 90 degrees, ρ 22i < ρ 22o . Therefore, the Hertzian contact area on the inner race side is larger than the Hertzian contact area on the outer race side. The characteristics of the spin moment M si on the inner race side, the characteristics of the spin moment M so on the outer race side and the power loss characteristics calculated using the aforesaid elastic model will now be described referring to FIG. 6. The rotation axis angle γ of the ball 67a is determined by the balance between the inner race spin moment M si and the outer race spin moment M so . The spin moment M si of the inner race and spin moment M so of the outer race are balanced at the first rotation axis angle γ0 , second rotation axis angle γ1 , and third rotation axis angle γ2 . The power loss ( M si •ω si + M so •ω so ) is smallest at the rotation axis angle γ2 where the spin angular velocity ω si of the inner race is smaller than the spin angular velocity ω so of the outer race, and during the running of the toroidal CVT, the rotation axis angle γ2 at which power losses are minimized, is maintained. By simply setting the magnitudes of the pitch diameters d i , d o of the inner race groove 67c and outer race groove 67d, heating of the inner race 64 can be prevented. Next, the invention will be described referring to FIG. 7 and FIG. 8. Referring to FIG. 7, in a power roller according to the invention, the radius of curvature r i of the inner race groove 67c is smaller than the radius of curvature r o of the outer race groove 67d. Further, the pitch diameter d i of the inner race groove 67c is smaller than the pitch diameter d o of the outer race groove 67d. In this way, the Hertzian contact area between the inner race groove 67c and the ball 67a, is set larger than the Hertzian contact area between the outer race groove 67d and the ball 67a. As the radius of curvature r i of the inner race groove 67c is set smaller than the radius of curvature r o of the outer race groove 67d, the Hertzian contact area on the inner race side is set still larger than the case of FIG. 5. As shown in FIG. 8, the inner race spin moment M si is larger than the inner race spin moment M si of the case of FIG. 5. As a result, the difference of power loss between the state where the outer race spin moment is small (state of rotation axis angle γ1 ) and the state where the inner race spin moment is small (state of rotation axis angle γ2 ), increases, so the toroidal CVT can definitely be run in the state where the spin angular velocity ω si of the inner race is smaller than the spin angular velocity ω so of the outer race. Describing the effect, in the power roller according to the invention, an even larger difference in the Hertzian contact area can be produced between the inner race and outer race than in the case of FIG. 3. For this reason, maintenance of the rotation axis angle γ2 where the inner race spin is small, is more reliable. Therefore, decrease of power transmission performance due to the ball spin of the thrust ball bearing 27 can be prevented. As discussed above, as the rotation axis angle γ of the ball is determined by the balance between the inner race spin moment M si between the ball and inner race groove, and the outer race spin moment M so between the ball and outer race groove, the rotation axis angle γ of the ball is set to the angle γ2 at which the inner race spin angular velocity ω si is less than the outer race spin angular velocity ω so . Further, the above teaching may be applied also to an infinitely variable transmission (IVT) wherein a toroidal continuously variable transmission is used.

Claims: A power roller for a toroidal continuously variable transmission, comprising: an inner race (64) for transmitting power between an input disk (61) and an output disk (62) via an oil film, an outer race (66) supported by a power roller support member (65), and a thrust ball bearing (67) for supporting the inner race (64) on the outer race (66) such that the inner race (64) is free to rotate, wherein the thrust ball bearing (67) has an inner race groove (67c) having an arc-shaped cross-section formed in the inner race (64), and an outer race groove (67d) having an arc-shaped cross-section formed in the outer race (66), an annular space defined by the inner race groove (67c) and outer race groove (67d), plural balls (67a) disposed in the annular space, and a holder (67b) for supporting the balls, wherein a first spin occurs in the contact part between each of the balls (67a) and the inner race groove (67c), which is the slip between each of the balls (67a) and inner race groove (67c), wherein each of the balls (67a) and the inner race groove (67c) rotate relative to each other, and a second spin occurs in the contact part between each of the balls (67a) and the outer race groove (67d), which is the slip between each of the balls (67a) and the outer race groove (67d), wherein each of the balls (67a) and outer race groove (67d) rotate relative to each other, and wherein the inner race groove (67c) has a pitch diameter less than the pitch diameter of the outer race groove (67d) such that the direction of the rotation axis (50) of the balls (67a), defined by the balance between the frictional moment of the first spin and the frictional moment of the second spin, is set to a direction in which the angular velocity of the first spin is less than the angular velocity of the second spin, characterized in that the inner race groove (67c) has a radius of curvature less than the radius of curvature of the outer race groove (67d). Toroidal continuously variable transmission comprising at least one power roller according to claim 1.

Inventors: YAMAMOTO TAKESHI, YAMAMOTO, TAKESHI
Applicants: NISSAN MOTOR, NISSAN MOTOR COMPANY, LIMITED
Application year: 2001
Main classification: F16H15
IPC codes: F16H15, F16C33, F16C19