Abstract:
A continuously variable transmission apparatus includes: an input shaft, an output shaft, a toroidal continuously variable transmission, a gear-type differential unit including a plurality of gears, and a controller. The controller calculates a torque actually passing through the toroidal continuously variable transmission to obtain a deviation of the torque from a target value and adjusts a transmission ratio of the toroidal continuously variable transmission to eliminate the deviation. The controller stops the adjustment of the transmission ratio when the torque is not stable.

Description:
BACKGROUND OF THE INVENTION 
   1. Field of the Invention 
   The present invention relates to improvement of a continuously variable transmission apparatus, for use in an automatic transmission for a vehicle (automobile), in which a toroidal continuously variable transmission is incorporated. Particularly, the invention relates to a continuously variable transmission apparatus to which is added a function of improving characteristics during a vehicle stop or during very low speed driving while implementing a structure which avoids providing a driver with an uncomfortable feeling at vehicle start. 
   2. Background Art 
   A toroidal-type continuously variable transmission as shown in  FIGS. 7 to 9  has been investigated as an automatic transmission for a vehicle, and has found limited use. The toroidal continuously variable transmission is referred to as a double cavity type, in which input discs  2 ,  2  are supported on peripheries of both end portions of an input shaft  1  via ball splines  3 ,  3 . Accordingly, the two input discs  2 ,  2  are rotatably supported such that they are concentric with each other and rotate synchronously. Further, an output gear  4  is supported rotatably with respect to the input shaft  1  on the periphery of the intermediate portion of the input shaft  1 . Output discs  5 ,  5  are splined at respective end portions of a cylindrical portion provided at the center portion of the output gear  4 . Hence, the two output discs  5 ,  5  rotate synchronously with the output gear  4 . 
   A plurality of power rollers  6 ,  6  (usually two or three power rollers on each side) are interposed between the input discs  2 ,  2  and the output discs  5 ,  5 . The power rollers  6 ,  6  are rotatably supported on the inner surfaces of trunnions  7 ,  7  respectively via support shafts  8 ,  8  and a plurality of roller bearings. The trunnions  7 ,  7  are provided so as to swing around pivot shafts  9 ,  9  disposed for the respective trunnions  7 ,  7  on both end portions thereof, in longitudinal directions (i.e., in vertical directions in  FIGS. 7 and 9 , and in a direction perpendicular to the plane of  FIG. 8 ). The trunnions  7 ,  7  are inclined by hydraulic actuators  10 ,  10 ; specifically, the hydraulic actuators  10 ,  10  displace the trunnions  7 ,  7  along the axes of pivot shafts  9 ,  9 . Inclination angles of the trunnions  7 ,  7  are synchronized hydraulically and mechanically. 
   That is, in the case where the inclination angles of the trunnions  7 ,  7  are changed in order to change a transmission ratio between the input shaft  1  and the output gear  4 , the trunnions  7 ,  7  are displaced in opposite directions by the actuators  10 ,  10 , respectively. For example, the power roller  6  on the right-hand side in  FIG. 9  is displaced downward in  FIG. 9 , while the power roller  6  on the left-hand side in  FIG. 9  is displaced upward in  FIG. 9  by the same distance. As a result, forces acting along a tangential direction of the contact portions between the peripheral surfaces of the power rollers  6 ,  6  and the inner surfaces of the input side discs  2 ,  2  and the output side discs  5 ,  5  are changed in direction (in other words, sideslip occurs at contact portions thereof). Consequently, due to the change in direction of the forces, the trunnions  7 ,  7  swing (incline) in opposite directions around the pivot shafts  9 ,  9 , which are pivotally supported by support plates  11 ,  11 . As a result, contacting portions between the peripheral surfaces of the power rollers  6 ,  6  and the inner surfaces of the input discs  2 ,  2  and the output discs  5 ,  5  are changed. Thereby, a rotation transmission ratio between the input shaft  1  and the output gear  4  changes. 
   Pressurized oil is supplied to and discharged from the actuators  10 ,  10  by means of a single control valve  12 , irrespective of the number of the actuators  10 ,  10 . The movement of any one of the trunnions  7  is fed back to the control valve  12 . The control valve  12  has a sleeve  14  to be displaced in an axial direction (i.e., in the horizontal direction in  FIG. 9 , and in a direction perpendicular to the plane of  FIG. 7 ) by a stepping motor  13  and a spool  15  fitted into the inner periphery of the sleeve  14  so as to allow displacement in the axial direction thereof. A feedback mechanism is constituted as follows: rods  17 ,  17  connect the trunnions  7 ,  7  and pistons  16 ,  16  of the respective actuators  10 ,  10 ; and a precess cam  18  is fixed on an end portion of each of the rods  17  attached to any one of the trunnions  7 ,  7 . The movement of the rod  17 ; that is, a resultant total of the displacement in the axial direction and the displacement in the rotating direction, is transmitted to the spool  15  via the precess cam  18  and a link arm  19  to thereby displace the spool  15  in the axial direction. A synchronous cable  20  is suspended between the trunnions  7 ,  7  in such amanner that the inclination angles of the respective trunnions  7 ,  7  are mechanically synchronized to each other even in the case where trouble arises in a hydraulic system. 
   At the time of switching the transmission state, the sleeve  14  is displaced to a position corresponding to a desired transmission ratio by the stepping motor  13  to thereby open a flow path to a predetermined direction of the control valve  12 . As a result, the pressurized oil is supplied to the actuators  10 ,  10  in a predetermined direction, whereby the actuators  10 ,  10  displace the trunnions  7 ,  7  in a predetermined direction. That is, in accordance with supply of the pressurized oil, the trunnions  7 ,  7  swing around the pivot shafts  9 ,  9  while being displaced in the axial direction of the pivot shafts  9 ,  9 . Then, the motion (i.e., the motion in an axial direction and the swing) of one of the trunnions  7  is transmitted to the spool  15  via the precess cam  18  fixed to the end portion of the rod  17  and the link arm  19  to thereby displace the spool  15  in the axial direction. As a result, the flow path of the control valve  12  is closed in a state where the trunnions  7  are displaced by the predetermined amount, and supply and discharge of the pressurized oil to and from the actuators  10 ,  10  is stopped. 
   The operation of the control valve  12  based on the displacement of the trunnion  7  and the cam surface  21  of the precess cam  18  during the above is as follows. First, the trunnion  7  is displaced in the axial direction along with the opening of the flow path of the control valve  12 . Then, as described hitherto, in response to the sideslip generated on the contact portions between the peripheral surface of the power roller  6  and the inner peripheral surfaces of the input disc  2  and the output disc  5 , the trunnion  7  starts swinging around the pivot shaft  9 . Further, along with the displacement of the trunnion  7  in the axial direction, a displacement of a cam surface  21  is transmitted to the spool  15  via the link arm  19 . Thereby, the spool  15  is displaced in the axial direction and a state of the control valve  12  is changed. More specifically, the control valve  12  is switched, by the actuator  10 , in a direction of returning the trunnion  7  to a neutral position. 
   Accordingly, immediately after the displacement in the axial direction, the trunnion  7  starts displacement in the direction opposite that in which it has been displacing, toward the neutral position. However, the trunnion  7  continues swinging around the pivot shaft  9  as long as there exists a displacement from the neutral position. As a result, a displacement of the precess cam l 8  in a circumferential direction of a cam surface  21  is transmitted to the spool  15  via the link arm  19 , to thus displace the spool  15  in the axial direction. Then, under a state where the inclination angle of the trunnion  7  reaches a predetermined angle corresponding to the desired transmission ratio, the control valve  12  is closed simultaneously with the trunnion  7  returning to the neutral position. Hence, supply and discharge of the pressurized oil to and from the actuators  10 ,  10  is stopped. As a result, the inclination angle of the trunnion  7  becomes an angle corresponding to the amount of displacement of the sleeve  14  in the axial direction displaced by the stepping motor  13 . 
   During operation of the toroidal continuously variable transmission such that as described above, the input disc  2  (the left-hand input disc  2  in  FIGS. 7 and 8 ) is driven and rotated by a driving shaft  22  that is connected with a power source such as an engine via a hydraulic loader  23  as shown in  FIGS. 7 and 8 . As a result, the pair of the input discs  2 ,  2  supported on the respective end portions of the input shaft  1  rotate synchronously while being pressed in a direction approaching toward each other. Then, the rotational movement is transmitted to the output discs  5 ,  5 , via the power rollers  6 ,  6 , and output from the output gear  4 . 
   As described above, when the power is transmitted from the input discs  2 ,  2  to the output discs  5 ,  5 , a force is applied on the trunnions  7 ,  7  in a direction along the pivot shafts  9 ,  9 , which are provided on respective ends of the trunnions  7 ,  7 , due to friction on rolling contact portions (i.e., traction portions) between the peripheral surfaces of the power rollers  6 ,  6  supported on the inner surfaces, and the inner surfaces of the discs  2 ,  5 . This force is referred to as a 2Ft, and the magnitude of the force is proportional to a torque transmitted from each of the input side discs  2 ,  2  to each of the output side discs  5 ,  5  (or from the output side discs  5 ,  5  to the input side discs  2 ,  2 ). Such a force 2Ft is supported by the actuators  10 ,  10 . Therefore, during operation of the toroidal continuously variable transmission, a pressure differential between a pair of oil pressure chambers  24   a  and  24   b  provided on respective sides of the pistons  16 ,  16  constituting the actuators  10 ,  10  is proportional to the magnitude of the force 2Ft. 
   In the case where rotational speeds of the input shaft  1  and the output gear  4  are changed, when deceleration is performed between the input shaft  1  and the output gear  4  first, the trunnions  7 ,  7  are moved in the axial directions of the pivot shafts  9 ,  9  by the actuators  10 ,  10 , thereby swinging the trunnions  7 ,  7  to a position shown in  FIG. 8 . Then, as shown in  FIG. 8 , the peripheral surfaces of the power rollers  6 ,  6  abut against portions of the inner surfaces of the input discs  2 ,  2  near the center and portions of the inner surfaces  4   a  of the output discs  5 ,  5  near the outer periphery, respectively. In contrast, at the time of increasing the speed, the trunnions  7 ,  7  are made swung in the opposite direction to that shown in  FIG. 8 . Accordingly, the trunnions  7 ,  7  are inclined so that, in the reverse state of that shown in  FIG. 8 , the peripheral surfaces of the power rollers  6 ,  6  abut against areas located slightly toward the outer peripheries of the inner surfaces of the input discs  2 ,  2  and areas located slightly toward the centers of the inner surface of the output discs  5 ,  5 , respectively. When the inclination angle of the trunnions  7 ,  7  is set at an intermediate angle between the above two angles, an intermediate transmission ratio (speed ratio) can be obtained between the input shaft  1  and the output gear  4 . 
   Further, for the case where the toroidal continuously variable transmission which is constituted and functions as described above is incorporated into an actual continuously variable transmission for a vehicle, there has been previously proposed combining the transmission with a differential unit such as a planetary gear mechanism to thereby constitute a continuously variable transmission apparatus. For example, U.S. Pat. No. 6,251,039 discloses a so-called geared-neutral-type continuously variable transmission apparatus which can switch rotation of an output shaft between forward and reverse with a stop state interposed therebetween while an input shaft rotates in a single direction.  FIG. 10  shows the continuously variable transmission apparatus disclosed in U.S. Pat. No. 6,251,039. The continuously variable transmission apparatus is constituted by combining a toroidal continuously variable transmission  25  and a planetary-gear-type transmission  26 . The toroidal continuously variable transmission  25  is provided with an input shaft  1 , a pair of input discs  2 ,  2 , an output disc  5   a , and a plurality of power rollers  6 ,  6 . In the example shown in  FIG. 10 , the output discs  5   a  is constituted such that outer surfaces of the pair of output discs abut each other to be formed integrally. 
   The planetary-gear-type transmission  26  is provided with a carrier  27  which is fixedly connected on the input shaft  1  and one of the input discs  2  (the right-hand input disc in  FIG. 10 ). A first transmission shaft  29  is rotatably supported on an intermediate portion in the radial direction of the carrier  27 , and planetary gear elements  28   a ,  28   b  are fixedly disposed on respective end portions of the first transmission shaft  29 . Further, a second transmission shaft  31  is rotatably supported, with the carrier  27  disposed between the second transmission shaft  31  and the input shaft  1  concentrically with the input shaft  1 , and sun gears  30   a ,  30   b  fixedly disposed on respective ends of the second transmission shaft  31 . Furthermore, the planetary gear elements  28   a  is meshed with a sun gear  33  which is fixedly disposed on a tip portion of a hollow rotary shaft  32  whose base portion (the left end portion in  FIG. 10 ) is connected to the output disc  5   a , and/or the planetary gear element  28   b  is meshed with the sun gear  30   a  which is fixedly disposed on one end portion (the left end portion in  FIG. 10 ) of the second transmission shaft  31 . The planetary gear element  28   a  (the left-hand element in  FIG. 10 ) is also meshed with a ring gear  35  which is rotatably provided around the carrier  27  via another planetary gear element  34 . 
   Meanwhile, planetary gear elements  37   a ,  37   b  are rotatably supported on a second carrier  36  which is provided around the sun gear  30   b  which is fixedly disposed on the other end portion (the right end portion in  FIG. 10 ) of the second transmission  31 . The second carrier  36  is fixedly disposed on a base end portion (the left end portion in  FIG. 10 ) of an output shaft  38  which is provided concentrically with the input shaft  1  and the second transmission shaft  31 . The planetary gear elements  37   a ,  37   b  are meshed with each other. Further, the planetary gear element  37   a , one of the planetary gear elements, is meshed with the sun gear  30   b . The other planetary gear element  37   b  is meshed with a second ring gear  39  which is rotatably provided around the second carrier  36 . The ring gear  35  and the second carrier  36  are allowed to engage and disengage by way of a low-speed clutch  40 . The second ring gear  39  and a stationary portion such as a housing are allowed to engage and disengage by way of a high-speed clutch  41 . 
   In the case of the continuously variable transmission apparatus shown in  FIG. 10  described above, under a so-called low-speed mode where the high-speed clutch  41  is disengaged simultaneously with engagement of the low-speed clutch  40 , the power of the input shaft  1  in transmitted to the output shaft  38  via the ring gear  35 . By changing the transmission ratio of the toroidal continuously variable transmission  25 , the overall transmission ratio of the continuously variable transmission apparatus; that is, the transmission ratio between the input shaft  1  and the output shaft  38 , is changed. Under such a low-speed mode, the overall transmission ratio of the continuously variable transmission apparatus changes infinitely. In other words, by adjusting the transmission ratio of the toroidal continuously variable transmission  25 , a rotation state of the output shaft can be switched between forward and reverse with a stop state interposed therebetween while the input shaft rotates in a single direction. 
   During acceleration or during constant-speed driving under such a low speed mode, a torque passing through the toroidal continuously variable transmission  25  (hereinafter referred to as “passing torque”) is applied on the output disc  5   a  from the input shaft  1  via the carrier  27 , the first transmission shaft  29 , the sun gear  33 , and the hollow rotary shaft  32 . Further, the torque is applied on the input discs  2 ,  2  from the output disc  5   a  via the power rollers  6 ,  6 . In other words, the torque passing through the toroidal continuously variable transmission  25  is circulated in a direction where the input discs  2 ,  2  receive torque from the power rollers  6 ,  6  during acceleration or constant-speed driving. 
   In contrast to the above, under a so-called high-speed mode where the low-speed clutch  40  is disengaged and the high-speed clutch  41  is engaged, the power of the input shaft  1  is transmitted to the output shaft  38  via the first and second transmission shafts  29 ,  31 . By changing the transmission ratio of the toroidal continuously variable transmission  25 , the overall transmission ratio of the continuously variable transmission apparatus is changed. In the above case, the higher the transmission ratio of the toroidal continuously variable transmission  25 , the higher the overall transmission ratio of the continuously variable transmission apparatus. 
   Note that the torque passing through the toroidal continuously variable transmission  25 —during acceleration or constant—speed driving under such a high speed mode—is applied in a direction where the input discs  2 ,  2  add torque on the power rollers  6 ,  6 . 
   For example, in the case of a continuously variable transmission apparatus having such a structure as shown in  FIG. 10  and capable of implementing a so-called infinitely variable transmission ratio where the output shaft  38  is stopped while the input shaft  1  rotates, it is important to maintain a torque applied on the toroidal continuously variable transmission  25  at an appropriate value under a state where the output shaft  38  is stopped and the transmission ratio is drastically increased, in view of ensuring durability and easy operability of the toroidal continuously variable transmission  25 . The reason for the above is as follows. As is clear from a relation of “rotational driving power=rotation speed×torque,” under a state where the transmission ratio is extremely high and the output shaft  38  stops, or rotates at a very low speed, with the input shaft  1  rotating, the torque passing through the toroidal continuously variable transmission  25  (passing torque) becomes larger than the torque applied on the input shaft  1 . Therefore, in order to secure durability of the toroidal continuously variable transmission  25  without upsizing the toroidal continuously variable transmission  25 , there must be adopted strict control for confining the torque within a range of appropriate values. More specifically, a control inclusive of a driving source is required for stopping the output shaft  38  while minimizing a torque input onto the input shaft  1 . 
   Meanwhile, under a state where the transmission ratio is extremely high, the torque applied on the output shaft  38  changes to a large extent even when the transmission ratio of the toroidal continuously variable transmission  25  is changed slightly. Accordingly, unless the transmission ratio of the toroidal continuously variable transmission  25  is adjusted strictly, the driver may experience an uncomfortable feeling, or drivability may be poor. For example, in the case of an automatic transmission for a vehicle, a stopping state is sometimes maintained while the driver steps on the brake during a vehicle stop. Under such a state, when the transmission ratio of the toroidal continuously variable transmission  25  is not adjusted strictly and a large torque is applied on the output shaft  38 , a force required for stepping on the brake pedal during the vehicle stop becomes larger, thereby increasing driver fatigue. Meanwhile, when the transmission ratio of the toroidal continuously variable transmission  25  is not adjusted strictly and too small a torque is applied on the output shaft  38 , vehicle start may fail to be smooth, or the vehicle may roll in reverse while starting on an uphill grade. Therefore, during a vehicle stop or very low speed driving, strict adjustment of the transmission ratio of the toroidal continuously variable transmission  25  is required in addition to control of the torque transmitted from the driving source to the input shaft  1 . 
   In consideration of the above, JP-A-10-103461 discloses a structure where the torque passing through the toroidal continuously variable transmission (passing torque) is regulated directly through control of a pressure differential between hydraulic actuator which are used for displacing trunnions. 
   However, in the case of the structure disclosed in JP-A-10-103461, because the control relies on only a pressure differential, stopping a posture of the trunnion at the moment when the passing torque has reached the desired value is difficult. More specifically, because a displacement amount of the trunnion becomes large in order to control the torque, there easily occurs so-called overshoot (and hunting resulting from the overshoot). . . where the trunnion continues displacement without stopping at the moment when the passing torque becomes coincident with the target value. Hence, control of the passing torque is not stable. 
   In particular, the overshoot is easily introduced in a case of a toroidal continuously variable transmission  25  of a so-called “without cast angle type,” as in the case of general half-toroidal continuously variable transmissions shown in  FIGS. 7 through 9 , wherein a direction along the pivot shafts  9 ,  9  which are provided on both end portions of the trunnions  7 ,  7  and a direction of center shafts of the input and output discs  2 ,  5  are perpendicular to each other. In contrast, in a case of a continuously variable transmission whose structure includes a cast angle as in the case of a general full-toroidal continuously variable transmission, a force in a direction of converging an overshoot acts thereon. Therefore, sufficient torque control is conceivably performed even with the structure disclosed in the above-cited JP-A-10-103461. 
   In view of the above circumstances, there can be conceived a control method or a control device with which a torque passing through the toroidal continuously variable transmission (passing torque) can be controlled strictly even in the case of a continuously variable transmission apparatus including a toroidal continuously variable transmission without cast angle, as in the case of a general half-toroidal continuously variable transmission. 
     FIG. 11  shows an example structure of a continuously variable transmission apparatus having such a control method and a control device as described above. The continuously variable transmission apparatus shown in  FIG. 11  has a function similar to that of a conventionally known continuously variable transmission apparatus shown in the above-cited  FIG. 10 ; however, assembly of the planetary-gear-type transmission  26   a  portion is improved by contriving a structure of the planetary-gear-type transmission  26   a  portion. 
   First and second planetary gears  42 ,  43 , both being of double pinion type, are supported on respective sides of a carrier  27   a  which is rotated together with the input shaft  1  and the pair of input discs  2 ,  2 . That is, the first planetary gear  42  is constituted of a pair of planetary gear elements  44   a ,  44   b , and the second planetary gear  43  is constituted of a pair of planetary gear elements  45   a ,  45   b . The planetary gear elements  44   a ,  44   b  are meshed with each other, as are the planetary gear elements  45   a ,  45   b . Further, the planetary gears elements  44   a ,  45   a  on the inner periphery are meshed with first and second sun gears  47 ,  48 , respectively, which are fixedly disposed on a tip portion (the right end portion in  FIG. 11 ) of a hollow rotary shaft  32   a  whose base portion (the left end portion in  FIG. 11 ) is connected to the output disc  5   a , and on one end portion (the left end portion in  FIG. 11 ) of a transmission shaft  46 . The planetary gear elements  44   b ,  45   b  on the outer periphery are meshed with a ring gear  49 . 
   Meanwhile, planetary gear elements  51   a ,  51   b  are rotatably supported on a second carrier  36   a  which is provided around a third sun gear  50 —which is fixedly disposed on the other end portion (the right end portion in  FIG. 11 ) of the transmission shaft  46 . The second carrier  36   a  is fixedly disposed on the base end portion (the left end portion in  FIG. 11 ) of an output shaft  38   a , which is concentrically provided with the input shaft  1 . The planetary gear elements  51   a ,  51   b  are meshed with each other. Further, the planetary gear element  51   a  on the inner periphery side is meshed with the third sun gear  50 , and the planetary gear element  51   b  on the outer periphery side is meshed with a second ring gear  39   a , which is rotatably provided around the second carrier  36   a . The ring gear  49  and the second carrier  36   a  are allowed to engage and disengage by way of a low-speed clutch  40   a . The second ring gear  39   a  and a stationary portion such as a housing are allowed to engage and disengage by way of a high-speed clutch  41   a.    
   In the case of an improved continuously variable transmission apparatus structured as described above, under a state where the high-speed clutch  41   a  is disengaged simultaneously with engagement of the low-speed clutch  40   a,  the power of the input shaft  1  is transmitted to the output shaft  38   a  via the ring gear  49 . By changing the transmission ratio of the toroidal continuously variable transmission  25 , an overall speed ratio e CVT  of the continuously variable transmission apparatus; that is, a speed ratio between the input shaft  1  and the output shaft  38   a , is changed. In the above case, a relationship between the speed ratio (i.e., the transmission ratio) e CVU  of the toroidal continuously variable transmission  25  and the overall speed ratio e CVT  of the continuously variable transmission apparatus, where a ratio between a number of teeth m 49  of the ring gear  49  and a number of teeth m 47  of the first sun gear  47  is set at i 1  (=m 49 /m 47 ), can be represented by the following Equation 1.
 
 e   CVT =( e   CVU   +i   1 −1)/i 1   (1)
 
   In the case where, for example, the ratio i 1 , the numbers of teeth, is 2, the relationship between the two speed ratios of e CVU  and e CVT  changes as shown by a line α in  FIG. 12 . 
   Meanwhile, under a state where the low-speed clutch  40   a  is disengaged and the high-speed clutch  41   a  is engaged, the power of the input shaft  1  is transmitted to the output shaft  38   a  via the first planetary gear  42 , the ring gear  49 , the second planetary gear  43 , the transmission shaft  46 , the planetary gear elements  51   a ,  51   b , and the second carrier  36   a . By changing the speed ratio e CVU  of the toroidal continuously variable transmission  25 , the overall speed ratio e CVT  of the continuously variable transmission apparatus is changed. A relationship in the above case between the speed ratio e CVU  of the toroidal continuously variable transmission  25  and the overall speed ratio e CVT  of the continuously variable transmission apparatus can be represented by the following Equation 2. In Equation 2, i 1  represents a ratio (=m 49 /m 47 ) between the number of teeth m 49  of the ring gear  49  and the number of teeth m 47  of the first sun gear  47 , i 2  represents a ratio (=m 49 /m 48 ) between the number of teeth m 49  of the ring gear  49  and a number of teeth m 48  of the second sun gear  48 , and i 3  represents a ratio between a number of teeth m 39  of the second ring gear  39   a  and a number of teeth m 50  of the third sun gear  50  (=m 39 /m 50 ).
 
 e   CVT   ={l/ (1 −i   3 )}·{ l+ ( i   2   /i   1 )( e   CVU −1)}  (2)
 
   In the case where the ratio i 1  is 2, i 2  is 2.2, and i 3  is 2.8, the relationship between the two speed ratios of e CVU  and e CVT  changes as shown by a line β in  FIG. 12 . 
   As is clear from the line α in  FIG. 12 , a continuously variable transmission apparatus which is constituted and functions in the aforesaid manner can realize a so-called infinitely variable transmission ratio state where the output shaft  38   a  is stopped while the input shaft  1  rotates. However, as mentioned previously, under such a state where the output shaft  38  is stopped or driven at a very low speed while the input shaft  1  rotates, the torque passing through the toroidal continuously variable transmission  25  (i.e., passing torque) becomes greater than a torque applied on the input shaft  1  from the engine which is the driving source. For this reason, the torque which is input from the driving source into the input shaft  1  must be regulated properly during a vehicle stop or during very low speed driving in order to prevent the passing torque from becoming excessively large (or excessively small). 
   Further, during the very low speed driving, under a state where the output shaft  38   a  is almost stopped; in other words, under a state where the transmission ratio of the continuously variable transmission apparatus is significantly large and the rotation speed of the output shaft  38   a  is significantly slower than that of the input shaft  1 , the torque applied on the output shaft  38   a  fluctuates to a large extent upon a slight fluctuation in the transmission ratio of the continuously variable transmission apparatus. Therefore, the torque which is input also from the driving source to the input shaft  1  must be regulated properly in order to secure smooth drivability. 
   During acceleration or constant-speed driving under such a low-speed mode, the torque is, as is the case with the conventional structure shown in the aforementioned  FIG. 10 , applied on the output disc  5   a  from the input shaft  1  via the carrier  27   a , the first planetary gear  42 , the first sun gear  47 , and the hollow rotary shaft  32   a . Further, the torque is applied on the input discs  2 ,  2  from the output disc  5   a  via the power rollers  6 ,  6  (see  FIG. 10 ). In other words, the passing torque is circulated in a direction where the input discs  2 ,  2  receive torque from the power rollers  6 ,  6  during acceleration or constant-speed driving. 
   For this reason, as shown in  FIG. 13 , a control method and a control apparatus according to the above constitution are arranged such that the torque input from the driving source into the input shaft  1  is regulated properly. First, a rotation speed of the engine which serves as a driving source is controlled roughly. Specifically, the rotation speed of the engine is regulated to a point “a” in the range of “w” of  FIG. 13 . In conjunction with the above, there is set the transmission ratio of the toroidal continuously variable transmission  25  which is required for matching a rotation speed of the input shaft  1  of the continuously variable transmission apparatus with the controlled rotation speed of the engine. This setting is to be operated according to the above-mentioned Equation 1. That is, the torque transmitted from the engine to the input shaft  1  must be strictly regulated in the case of a so-called low-speed mode where the low-speed clutch  40   a  is engaged and the high-speed clutch  41   a  is disengaged. Therefore, the transmission ratio of the toroidal continuously variable transmission  25  is to be set, according to Equation 1, such that the rotation speed of the input shaft  1  corresponds to the required rotation speed of the output shaft  38   a.    
   Meanwhile, a pressure differential between the oil pressure chambers  24   a ,  24   b  (see  FIG. 9  and  FIG. 15  described later) incorporated in the hydraulic actuators  10 ,  10 —used for displacing the trunnions  7 ,  7  incorporated in the toroidal continuously variable transmission  25  in the direction along the pivot shafts  9 ,  9 —is measured with an oil pressure sensor  52  (see  FIG. 2 , described later). The oil pressure is measured under a state where the rotation speed of the engine is roughly controlled (however, the rotation speed must be maintained constant) and, corresponding thereto, the transmission ratio of the toroidal continuously variable transmission  25  is set according to Equation 1 in the manner described above. Then, the torque passing through the toroidal continuously variable transmission  25  (passing torque) T CVU  is calculated from the oil pressure differential obtained from the measurement. 
   Specifically, so long as the transmission ratio of the toroidal continuously variable transmission  25  is constant, the oil pressure differential is proportional to the torque T CVU  passing through the toroidal continuously variable transmission  25 . Accordingly, the torque T CVU  can be calculated from the above oil pressure differential. The reason for this is as follows. As described above, the actuators  10 ,  10  support a force of so-called 2Ft having a magnitude proportional to the torque (i.e., the torque T CVU  passing through the toroidal continuously variable transmission  25 ) transmitted from the input discs  2 ,  2  to the output disc  5   a  (or from the output discs  5   a  to the input discs  2 ,  2 ). 
   Meanwhile, the torque T CVU  can be obtained from Equation 3 below.
 
 T   CVU   =e   CVU   ·T   IN   /{e   CVU +( i   1 −1)η CVU }  (3)
 
   In Equation 3, e CVU  represents a speed ratio of the toroidal continuously variable transmission  25 , T IN  represents the torque input from the engine to the input shaft  1 , i 1  represents a teeth number ratio of planetary-gear-type transmission pertaining to the first planetary gear  42  (i.e., a ratio between the number of teeth m 49  of the ring gear  49  and the number of teeth m 47  of the first sun gear  47 ), and η CVU  represents efficiency of the toroidal continuously variable transmission  25 . 
   Here, a deviation ΔT (=T CVU1 −T CVU2 ) is obtained from T CVU1 , which is the torque actually passing through the toroidal continuously variable transmission  25  as obtained from the above oil pressure differential and the target passing torque T CVU2  obtained from Equation 3. Then, the speed ratio of the toroidal continuously variable transmission  25  is adjusted in a direction where the deviation ΔT is eliminated (i.e., where ΔT becomes zero). Note that because the deviation of the torque ΔT and a deviation of the oil pressure differential are in a proportional relationship, the adjustment of the transmission ratio can be performed either by the deviation of the torque or by the deviation of the oil pressure differential. In other words, from the technical point of view, control of the transmission ratio based on the deviation of the torque is identical with control of the transmission ratio based on the deviation of the oil pressure differential. 
   As an example, the following is conceived under the assumption that, within the range where the actual torque T CVU1  (measured value) passing through the toroidal continuously variable transmission  25  is restricted to the target value T CVU2  as shown in  FIG. 13 , a torque T IN  with which the engine drives the input shaft  1  changes in such a sharply decreasing direction that the rotation speed of the input shaft  1  is increased. Such characteristics of the engine can be easily obtained even in a low-speed rotation range when the engine is electronically controlled. In a case where the engine has such characteristics and where the measured torque value T CVU1  has a deviation from the target torque value T CVU2  in the direction in which the input discs  2 ,  2  receive torque from the power rollers  6 ,  6  (see  FIGS. 8 through 10 ), the overall transmission ratio of the continuously variable transmission apparatus is displaced to the deceleration side so as to increase the rotation speed of the engine to thereby reduce the torque T IN  which drives the input shaft  1 . To achieve the above, the transmission ratio of the toroidal continuously variable transmission  25  is changed to the acceleration side. 
   However, under a vehicle stop state where the driver steps on a brake pedal (i.e., a state where the rotation speed of the output shaft is zero), the transmission ratio of the toroidal continuously variable transmission  25  is controlled within a range where the adjusted force can be absorbed by a slip generated in the toroidal continuously variable transmission  25 ; that is, a slip (creep) generated on the contact portions (i.e., traction portion) of the inner surfaces of the input and output discs  2 ,  5   a  and the peripheral surfaces of the power rollers  6 ,  6  (see  FIGS. 8 through 10 ). Therefore, an allowable range for adjusting the speed ratio is limited to a range where strain is not applied on the contact portions, which is a stricter limitation than that imposed in the case of low-speed driving. 
   For example, when the target value T CVU2  is at point “a” and the measured value T CVU1  is at point “b” in  FIG. 13 , the input discs  2 ,  2  have deviation in a direction receiving a torque from the power rollers  6 ,  6 . Here, the speed ratio e CVU  of the toroidal continuously variable transmission  25  is changed to the acceleration side so that the overall speed ratio e CVT  of the continuously variable transmission apparatus (T/M) is changed to the deceleration side. A rotation speed of the engine is increased in conjunction with the above so as to lower the torque. In contrast, when the measured value T CVU1  is at point “c” in  FIG. 13 , the input discs  2 ,  2  have deviation in a direction where torque is added on the power rollers  6 ,  6 . In the case where T CVU1  is at point “c,” reverse to the case where T CVU1  is at the point “b,” the speed ratio e CVU  of the toroidal continuously variable transmission  25  is changed to the deceleration side so that the overall speed ratio e CVT  of the continuously variable transmission apparatus (T/M) is changed to the acceleration side. The rotation speed of the engine is decreased in conjunction with the above so as to increase the torque. 
   The above-mentioned operations are repeated until the torque T CVU1  actually passing through the toroidal continuously variable transmission  25  as obtained from the oil pressure differential matches the target value. In other words, the above-mentioned operations are repeated in the case where the torque T CVU1  passing through the toroidal continuously variable transmission  25  cannot be matched with the target value T CU2  through only one iteration of transmission gear control of the toroidal continuously variable transmission  25 . As a result, the torque T IN  with which the engine rotates and drives the input shaft  1  can be set closer to a value which allows the torque T CVU  passing through the toroidal continuously variable transmission  25  to reach the target value T CVU2 . 
   Note that the above operations are performed automatically and in a short period of time through instructions from a microcomputer which is incorporated in a controller of the continuously variable transmission apparatus. 
     FIG. 14  shows relationships among a ratio (the left-handed vertical axis) of the torque T CVU  passing through the toroidal continuously variable transmission  25  and the torque T IN  with which the engine rotates and drives the input shaft  1 , an overall speed ratio e CVT  (horizontal axis) of the continuously variable transmission apparatus, and a speed ratio e CVU  (the right-handed vertical axis) of the toroidal continuously variable transmission  25 . The solid line “a” shows a relationship between the ratio of the passing torque T CVY  to the driving torque T IN  and the overall speed ratio e CVT  of the continuously variable transmission apparatus, and the dotted line “b” shows a relationship between the two speed ratios e CVT  and e CVU . The above constitution regulates the speed ratio e CVU  of the toroidal continuously variable transmission  25  so as to regulate the torque T CVU1  actually passing through the toroidal continuously variable transmission  25  to the target value (T CVU2 ) represented by points on the solid line “a” under a state where the overall speed ratio e CVT  of the continuously variable transmission apparatus is regulated to a predetermined value. 
   In the above constitution, control for regulating the torque T CVU1  actually passing through the toroidal continuously variable transmission  25  to the point on the solid line “a,” which is the target value T CVU2 , is performed in two stages. Specifically, the rotation speed of the engine is roughly controlled to a specific rotation speed; that is, to a value assumed to provide the target value T CVU2 . Thereafter, the transmission ratio of the toroidal continuously variable transmission  25  is controlled in conjunction with the rotation speed thereof. For this reason, in contrast to the case of a conventional method, the torque T CVU1  actually passing through the toroidal continuously variable transmission  25  can be regulated to the target value T CVU2  without introducing an overshoot (and resultant hunting), or even when introduced, the overshoot is suppressed within such a level that would not raise any problems in practical use. 
   Note that, as described above, under a state of a vehicle stop with the driver stepping on a brake pedal, a driving force (torque) is applied on the output shaft  38   a  ( FIG. 10 ) based on the slip generated within the toroidal continuously variable transmission  25 . The magnitude of the torque may be set to a value which corresponds to a creep torque generated in a general automatic transmission provided with a torque converter. The reason for this is to avoid providing an uncomfortable feeling to a driver who is accustomed to operations of a general automatic transmission. In addition, a direction of the torque is determined by a position of a control lever provided at a driver&#39;s seat. When a forward direction (D range) is selected by the control lever, a torque of a forward direction is applied on the output shaft  38   a . When a reverse (R range) is selected, a torque of reverse direction is applied. 
   Next, a circuit in a section which controls the speed ratio of the toroidal continuously variable transmission  25  so that the torque T CVU1  actually passing through the toroidal continuously variable transmission  25  matches the target value T CVU2  will be described with reference to  FIG. 15 . By way of the control valve  12   a , pressurized oil can be supplied to and discharged from the pair of oil pressure chambers  24   a ,  24   b  included in the hydraulic actuators  10 ,  10  which are used for displacing the trunnions  7 ,  7  in the axial direction (in the vertical direction in  FIG. 15 ) of pivot shafts  9 ,  9  (see  FIG. 9 ). The sleeve  14  constituting the control valve  12  is allowed to displace in the axial direction by a stepping motor  13  via a link arm  54  and a rod  53 . The spool  15  constituting the control valve  12  is engaged with the trunnion  7  via the link arm  19 , the precess cam  18 , and the rod  17 . The spool  15  is allowed to displace in the axial direction in conjunction with a displacement in the axial direction and a swing of the trunnion  7 . The above constitution is principally the same as that of a conventionally known toroidal continuously variable transmission. 
   The above constitution is particularly arranged such that the sleeve  14 , driven by the stepping motor  13 , can also be driven by a hydraulic differential pressure cylinder  55 . Specifically, a tip portion of the rod  53  whose base end portion is connected to the sleeve  14  is pivotally supported by an intermediate portion of the link arm  54 . Further, pins provided at output portions of the stepping motor  13  or the differential pressure cylinder  55  are engaged with elongated holes provided on respective end portions of the link arm  54 . When one pin in the elongated hole provided on one of the two end portions of the link arm  54  is pushed or pulled, the other pin in the elongated hole on the other end portion serves as a pivot. According to such a constitution, the sleeve  14  can be displaced in the axial direction not only by the stepping motor  13  but also by the differential pressure cylinder  55 . The above constitution is arranged such that the speed ratio e CVU  of the toroidal continuously variable transmission  25  can be adjusted by a displacement of the sleeve  14  caused by the differential pressure cylinder  55  depending on the torque T CVU  passing through the toroidal continuously variable transmission  25 . 
   In order to achieve the above, the constitution is arranged such that different oil pressures can be induced into a pair of oil pressure chambers  56   a ,  56   b  provided in the differential pressure cylinder  55  via a correcting control valve  57 . Oil pressures introduced into the oil pressure chambers  56   a ,  56   b  are determined from a pressure differential ΔP between oil pressures P DOWN  and P UP  which act in the pair of oil pressure chambers  24   a ,  24   b  constituting the actuator  10 ; and a pressure differential ΔP O  between output pressures of a pair of solenoid valves  58   a  and  58   b  used for adjusting opening of the correcting control valve  57 . Specifically, opening and closing of the two solenoid valves  58   a ,  58   b  are calculated by an unillustrated control device (hereinafter referred to as “controller”), and controlled on the basis of an output signal output from the controller such that the pressure differential ΔP O  between the output pressures of the two solenoid valves  58   a  and  58   b  reaches a target pressure differential corresponding to the target torque T CVU2  of the toroidal continuously variable transmission  25 . Accordingly, the following forces act on a spool  59  constituting the correcting control valve  57 : a force corresponding to the pressure differential ΔP between oil pressures acting on the oil pressure chambers  24   a ,  24   b  of the actuator  10 ; and a pressure differential ΔP O —between the output pressures of the solenoid valves  58   a ,  58   b —which is the target pressure differential corresponding to the target torque T CVU2 ; that is, counterforce against ΔP. 
   In the case where the torque T CVU1  actually passing through the toroidal continuously variable transmission  25  is identical with the target torque T CVU2 ; that is, in the case where a difference ΔT between the passing torque T CVU1  and the target torque T CVU2  is zero, the force corresponding to the pressure differential ΔP between oil pressures acting on the oil pressure chambers  24   a ,  24   b  of the actuator  10  and the force corresponding to the pressure differential ΔP O  between the output pressures of the solenoid valves  58   a ,  58   b  are balanced. For this reason, the spool  59  constituting the correcting control valve  57  is brought into a neutral position, and the pressures acting on the oil pressure chambers  56   a ,  56   b  of the differential pressure cylinder  55  become equal to each other. Under the above state, a spool  60  of the differential pressure cylinder  55  is brought into a neutral position, and the speed ratio of the toroidal continuously variable transmission  25  remains unchanged (not corrected). 
   Meanwhile, when a difference arises between the torque T CVU1  actually passing through the toroidal continuously variable transmission  25  and the target torque T CVU2 , balance is lost between the force corresponding to the pressure differential ΔP between the oil pressures acting on the oil pressure chambers  24   a ,  24   b  of the actuator  10  and the force corresponding to the pressure differential ΔP O  between the output pressures of the solenoid valves  58   a ,  58   b . Then, according to a magnitude and direction of the difference ΔT between the passing torque T CVU1  and the target torque T CVU2 , the spool  59  constituting the correcting control valve  57  is displaced in the axial direction, to thus induce an oil pressure corresponding to the magnitude and direction of ΔT into the oil pressure chambers  56   a ,  56   b  of the differential pressure cylinder  55 . Then, the spool  60  of the differential pressure cylinder  55  is displaced in the axial direction, whereby the sleeve  14  constituting the control valve  12  is displaced in the axial direction. Consequently, the trunnions  7 ,  7  are displaced in the direction along the pivot shafts  9 ,  9  to thus change (correct) the speed ratio of the toroidal continuously variable transmission  25 . Note that the direction and the amount of the displacement of the speed ratio in relation to the above is the same as described with reference to the aforementioned  FIGS. 13 and 14 . The amount of displacement of the speed ratio; that is, the amount to be corrected (i.e., amount to be corrected in relation to the speed ratio) of the toroidal continuously variable transmission  25  in relation to the above is sufficiently small as compared with a speed ratio width of the toroidal continuously variable transmission  25 . For this reason, a stroke of the spool  60  of the differential pressure cylinder  55  is designed so as to be sufficiently smaller than a stroke of an output portion of the stepping motor  13 . 
   In the case where the conventional continuously variable transmission apparatus shown in  FIG. 10  or the structure shown in  FIG. 11  is employed as an actual automatic transformer of a vehicle, when a non-travel range is selected with a shift lever provided at a driver&#39;s seat, each of low-speed clutches  40 ,  40   a  and high-speed clutches  41   a ,  41   b  are arranged to be disengaged. Specifically, in the case where the shift lever is in a neutral range (N range) or in a parking range (P range)—each used for selecting a state where a vehicle is not allowed to drive—each of the clutches  40 ,  40   a ,  41   a , and  41   b  is disengaged. As a result, the torque passing through the toroidal continuously variable transmission  25  and the planetary-gear-type transmission  26 ,  26   a  becomes quite small (substantially zero). Accordingly, durability of the toroidal continuously variable transmission  25  and the planetary-gear-type transmission  26 ,  26   a  can be secured. 
   However, in such a state where each of the clutches  40 ,  40   a ,  41   a ,  41   b  is disengaged and the torque passing through the toroidal continuously variable transmission  25  becomes quite small, the correction of the speed ratio of the toroidal continuously variable transmission  25  according to  FIGS. 13 and 14  cannot be performed accurately. More specifically, the speed ratio may be corrected excessively because of a failure to control the transmission ratio (speed ratio) based on the torque passing through the toroidal continuously variable transmission  25 . Further, under such a state that the speed ratio is corrected excessively, when a travel range (drive range or a reverse range) is selected with a shift lever, an excessive torque may be applied on the output shaft  38 ,  38   a  at a moment when the low-speed clutch  40 ,  40   a  is engaged. When an excessive torque is applied on the output shaft  38 ,  38   a  in this manner, a driver may feel undesirably uncomfortable feeling at a vehicle start. 
   SUMMARY OF THE INVENTION 
   The present invention has been conceived in view of the above-mentioned problems, and provides a continuously variable transmission apparatus which does not provide a driver with an uncomfortable feeling, because unnecessary control of a transmission ratio is hindered. 
   The invention provides a continuously variable transmission apparatus, including: an input shaft rotated and driven by a driving source; an output shaft; a toroidal continuously variable transmission; a gear-type differential unit including a plurality of gears; and a controller for controlling change of a transmission ratio of the toroidal continuously variable transmission. The toroidal continuously variable transmission includes: an input disc rotated and driven by the input shaft, an output disc supported so as to be relatively rotatable with respect to the input disc, a plurality of power rollers interposed between the input disc and the output disc, a plurality of support members rotatably supporting the respective power rollers, and an actuator having a pair of oil pressure chambers and displacing the support members so as to change the transmission ratio between the input disc and the output disc. The differential unit includes: a first input portion rotated and driven by the input shaft together with the input disc, and a second input portion connected to the output disc. A rotational movement which is obtained in accordance with a speed differential between the first and the second input portions is taken out and transmitted to the output shaft. The controller performs the following functions (1) through (3): (1) a function of adjusting the transmission ratio of the toroidal continuously variable transmission so as to change a relative displacement speed of the plurality of gears contained in the planetary-gear-type transmission, thereby converting the rotational condition of the output shaft between a forward-rotating condition and a reverse-rotating condition through a non-rotational condition while the input shaft is kept rotating in a single direction; (2) a function of measuring a differential pressure between a pair of chambers constituting an actuator so as to calculate a torque actually passing through the toroidal continuously variable transmission; thereafter obtaining a deviation, from a target value, of the torque actually passing through the toroidal continuously variable transmission; and thereby adjusting the transmission ratio of the toroidal continuously variable transmission so that the torque passing through a toroidal continuously variable transmission attains a target value, and (3) a function of stopping the above function (2) in the case where the torque passing through the toroidal continuously variable transmission fails to become stable. 
   In the case of a continuously variable transmission apparatus constituted as described above, when a travel state is selected so as to stop a vehicle or to drive at a low speed, as in the case of the constitution of  FIG. 11  described hitherto, a transmission ratio (speed ratio) of a toroidal continuously variable transmission is adjusted so as to regulate a torque passing through the toroidal continuously variable transmission to thus prevent providing an uncomfortable feeling to a driver. Further, in the case where a non-travel state is selected, or immediately after that a travel state is selected from a non-travel state, when the torque passing through the toroidal continuously variable transmission  25   a  fails to become stable, control of the transmission ratio of the toroidal continuously variable transmission is stopped. Therefore, the transmission ratio of the toroidal continuously variable transmission is prevented from being excessively corrected, and the continuously variable transmission apparatus is prevented from providing an uncomfortable feeling to a driver because an excessive torque is transmitted to an output shaft when switched to a travel state. 

   
     BRIEF DESCRIPTION OF THE DRAWINGS 
     The present invention may be more readily described with reference to the accompanying drawings, in which: 
       FIG. 1  is a schematic sectional view of one-half of a continuously variable transmission according to a first embodiment of the present invention; 
       FIG. 2  is a block diagram of the continuously variable transmission according to the first embodiment of the invention; 
       FIG. 3  is a flow chart for explaining operations of the continuously variable transmission according to the first embodiment of the invention; 
       FIG. 4  is a hydraulic circuit diagram showing a mechanism for adjusting a transmission ratio of a toroidal continuously variable transmission incorporated in the continuously variable transmission of the first embodiment; 
       FIG. 5  is a flow chart for explaining operations of a second embodiment of the present invention; 
       FIG. 6  is a flow chart for explaining operations of a third embodiment of the invention; 
       FIG. 7  is a sectional view showing an example of a conventionally known toroidal continuously variable transmission; 
       FIG. 8  is a sectional view along line A—A in  FIG. 7 ; 
       FIG. 9  is a sectional view along line B—B in  FIG. 7 ; 
       FIG. 10  is a schematic sectional view showing an example of a conventionally known continuously variable transmission; 
       FIG. 11  is a schematic sectional view showing an example of a continuously variable transmission whose transmission ratio is controlled by an improved control device based on a conventional control device; 
       FIG. 12  is a diagram showing a relation between a transmission ratio of a toroidal continuously variable transmission incorporated in the continuously variable transmission and overall transmission ratio of the continuously variable transmission apparatus (T/M); 
       FIG. 13  is a diagrammatic showing a relationship between a rotation speed of an engine and a torque for explaining a condition where a transmission ratio is controlled by an improved control device based on a conventional control device; 
       FIG. 14  is a diagram showing a relation between a torque passing through a toroidal continuously variable transmission and a transmission ratio, and an overall transmission ratio of the continuously variable transmission apparatus; and 
       FIG. 15  is a hydraulic circuit diagram showing a mechanism for adjusting a transmission ratio of a toroidal continuously variable transmission constituting an improved continuously variable transmission apparatus based on a conventional mechanism. 
   

   DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS 
   There will now be described an embodiment which employs a continuously variable transmission according to the present invention wherein a torque applied to an input shaft from a driving source preferably varies in accordance with a rotation speed. In order to realize function (2), the rotation speed of the driving source is roughly controlled. Further, a transmission ratio of a toroidal continuously variable transmission is set to a value which is assumed to be required for matching the controlled rotation speed of the driving source with the rotation speed of the input shaft. 
   The embodiment is preferably provided with a clutch which transmits rotation movements when connected. The case where the torque passing through a toroidal continuously variable transmission specified in (3) fails to become stable corresponds to a case where the clutch is disengaged upon selection of a non-travel state. 
   Under the above condition, a non-travel state is selected during a travel state. Thereafter, function (2) is ceased until elapse of a predetermined period of time. 
   Alternatively, in the case where a pressure differential between a pair of oil pressure chambers is equal to a predetermined value or less, function (2) is ceased. 
   Further, a load-detecting device for detecting a load of the driving source is provided. After a travel state is selected during a non-travel state, function (2) is ceased until the load-detecting apparatus detects an increase in a load. 
   [First Embodiment] 
     FIGS. 1 through 4  show a first embodiment of the invention. As shown in  FIG. 1 , a continuously variable transmission is constituted by combining a toroidal continuously variable transmission  25   a  and a planetary-gear-type transmission  26   b  of a differential unit. The constitution of the continuously variable transmission is principally the same as that of the continuously variable transmission of aforementioned conventional constitution shown in  FIG. 10 , or the structure shown in  FIG. 11 . In the continuously variable transmission of the embodiment, a planetary gear element—which is longer in an axial direction among the planetary gear elements constituting first and second planetary gears  42 ,  43  of the planetary-gear-type transmission  26   b —is used as a planetary gear element  61  and provided on the outer side with respect to a radial direction. The planetary gear element  61  is meshed with planetary gear elements  44   a ,  45   a  provided on the inner side with respect to the radial direction. Further, a ring gear having a small width is used as a ring gear  49   a  which meshes with the planetary gear element  61 . According to the structure shown in  FIG. 1 , the continuously variable transmission apparatus is conceivably reduced in weight by reducing the axial length of the ring gear  49   a  whose diameter is a large, thereby reducing the volume thereof. Functions of the continuously variable transmission are the same as those of the continuously variable transmission of aforementioned conventional constitution shown in  FIG. 10 , or the structure shown in  FIG. 11 . 
   Next, operation of the continuously variable transmission apparatus of the first embodiment will be explained with reference to  FIG. 1 , a block diagram of  FIG. 2 , and a flow chart of  FIG. 3 . In  FIG. 2 , heavy arrows indicate power transmission paths, solid lines indicate hydraulic circuits, and dotted lines indicate electric circuits. Output of an engine  62  is input to an input shaft  1  via a damper  63 . The damper  63  serves as an elastic joint which smoothes rotation of the engine  62  and transmits it to the input shaft  1 . The present invention is characterized in that correction of transmission ratio of a toroidal continuously variable transmission  25   a  is ceased when a torque passing through the toroidal continuously variable transmission  25   a  is quite small or is unstable under a state where a clutch device  67  is disengaged or during an engaging process of the clutch  67 . Thereby a vehicle on which the continuously variable transmission is mounted is prevented from providing an uncomfortable feeling to a driver at start of the vehicle (i.e., immediately after the start of driving). The constitution of the continuously variable transmission is the same as that shown in  FIG. 1 . Therefore, to the extent possible, descriptions in relation to  FIG. 2 , descriptions employ the same reference numerals as those in  FIG. 1  for equivalent elements. The invention is also characterized in that a torque fluctuation at a time of engagement of a low-speed clutch  40   a  is suppressed. Control performed during driving with a high-speed clutch  41   a  engaged does not fall within the scope of the present invention. 
   Power transmitted to the input shaft  1  is transmitted to input discs  2 ,  2  via a hydraulic loader  23   a  which constitutes the toroidal continuously variable transmission  25   a . Further, the power is transmitted to an output disc  5   a  via power rollers  6 . With regard to the discs  2 ,  5   a , rotation speed of the input discs  2  is measured by an input-side rotation speed sensor  64 , and that of the output disc  5   a  is measured by an output-side rotation speed sensor  65 . Then, measured values are input to a controller  66 . Accordingly, a transmission ratio (i.e., speed ratio) between the discs  2 ,  5   a  (of the toroidal continuously variable transmission  25   a ) can be calculated. The power transmitted to the input shaft  1  is further transmitted to a planetary-gear-type transmission  26   b  of a differential unit directly or via the toroidal continuously variable transmission  25   a . Then, a differential component of constitution members of the planetary-gear-type transmission  26   b  is taken out to an output shaft  38   a  via a clutch device  67 . Note that the clutch device  67  represents the low-speed clutch  40   a  and the high-speed clutch  41   a  shown in the aforementioned  FIG. 1  and in  FIG. 4  described later. In addition, in the example, the rotation speed of the output shaft  38   a  can also be detected by an output shaft rotation speed sensor  68 . Note that the output shaft rotation speed sensor  68  is provided as a fail-safe device for detecting occurrence of a failure of the input-side rotation speed sensor  64  and the output-side rotation speed sensor  65 . 
   Meanwhile, an oil pump  69  is driven by the power taken out from the damper  63 . Pressurized oil discharged from the oil pump  69  can be supplied to the hydraulic loader  23   a ; and to a control valve device  70  for controlling an amount of displacement of the actuator  10  (see  FIGS. 4 ,  9 , and  15 ) for displacing trunnions  7  which support the power rollers  6 . Note that the control valve device  70  is constituted of a control valve  12  shown in the aforementioned  FIG. 15 ; a differential cylinder  55 ; a correcting control valve  57 ; and a high-speed switch valve  71  and a low-speed switch valve  72  shown in  FIG. 4  to be described later. Oil pressures of a pair of oil pressure chambers  24   a ,  24   b  (see  FIGS. 4 ,  9 , and  15 ) provided in the actuator  10  are detected by an oil pressure sensor  52  (actually constituted by a pair of oil pressure sensors), and the detection signals are input to the controller  66 . The controller  66  calculates a passing torque of the toroidal continuously variable transmission  25   a  on the basis of the signal from the oil pressure sensor  52 . 
   Meanwhile, an operational status of the control valve device  70  can be switched by a stepping motor  13 ; a line-pressure-control solenoid valve  73 ; a solenoid valve  58   a  ( 58   b ) for switching the correcting control valve  57 ; and a shift solenoid valve  74  for switching the high-speed switch valve  71  and the low-speed switch valve  72 . Further, any of the stepping motor  13 , the line-pressure-control solenoid valve  73 , the solenoid valve  58   a  ( 58   b ) for switching the correcting control valve  57 , and the shift solenoid valve  74  can be switched in accordance with a control signal from the controller  66 . 
   In addition to signals from the rotation speed sensors  64 ,  65 , and  68  and the oil pressure sensor  52 , a detection signal of an oil temperature sensor  75 , a position signal of a position switch  76 , a detection signal of an accelerator sensor  77 , and a signal of a brake switch  78  are input to the controller  66 . The oil temperature sensor  75  detects a temperature of lubricant (i.e., traction oil) in a casing where the continuously variable transmission is contained. The position switch  76  is for generating a signal which indicates a position of a shift lever, which is provided at a driver&#39;s seat for the purpose of switching an oil-pressure-manual-switch valve  79  shown in  FIG. 4  described later. The accelerator sensor  77  detects a position of an accelerator pedal. The brake switch  78  detects that a brake pedal is being pressed or that a parking brake is actuated and generates a signal indicating the same. 
   The controller  66  sends control signals corresponding to signals from the switches  76 ,  78  and the sensors  52 ,  64 ,  65 ,  68 ,  75 ,  77  to the stepping motor  13 , the line-pressure-control solenoid valve  73 , the solenoid valve  58   a  ( 58   b ), and the shift solenoid valve  74 . In addition, the controller  66  sends control signals to an engine controller  80  for controlling the engine  62 . As is the case with the constitution of  FIG. 11  described hitherto, a speed ratio between the input shaft  1  and the output shaft  38   a  is changed, or a torque applied on the output shaft  38   a  passing through the toroidal continuously variable transmission  25   a  (passing torque) is controlled during a vehicle stop or during very low speed driving. 
   In order to implement the invention with the above constitution, when a non-travel state (i.e., P range or N range) is selected on the basis of a signal from the position switch  76 , the controller  66  of the continuously variable transmission according to the example disengages the low-speed clutch  40   a  and the high-speed clutch  41   a . Under the above condition, rotational movement of the input shaft  1  is not transmitted to the output shaft  38   a , irrespective of a transmission ratio (i.e., speed ratio) of the toroidal continuously variable transmission  25   a . Furthermore, under the above condition, the torque passing through the toroidal continuously variable transmission  25   a  is substantially zero (except for a torque corresponding to a quite small friction resistance). 
   Under such a condition, when a transmission ratio of the toroidal continuously variable transmission  25   a  is controlled on the basis of a torque passing through the toroidal continuously variable transmission  25   a  as described hitherto, the transmission ratio may be corrected excessively. In the example, in order to solve the problem, when non-travel is selected under a vehicle stop state, correction of the transmission ratio of the toroidal continuously variable transmission  25   a  by the correcting control valve  57  is ceased. Expressed another way, in the example, the correction of the transmission ratio of the toroidal continuously variable transmission  25   a  by the correcting control valve  57  under a state where a vehicle is stopped is not implemented until a travel state is selected and the torque passing through the toroidal continuously variable transmission  25   a  becomes stable. This will now be described with reference to  FIG. 3 . 
   Under a state where an automobile (vehicle) is stopped (Step  1 ), when a travel (D range or R range) state is judged to have been selected during a non-travel (P range or N range) state on the basis of a signal from the position switch  76  (Step  2 ), engagement of a clutch device  67  (the low-speed clutch  40   a  or the high-speed clutch  41   a ) is started (Step  3 ). Then, concurrently with the above, the stepping motor  13  (see  FIGS. 4 ,  9 , and  15 ) is positioned so that the transmission condition of the continuously variable transmission is set to establish a infinitely variable transmission ratio state where the output shaft  38   a  is stopped while the input shaft  1  rotates (Step  4 ). Note that the positioning of the stepping motor  13  for implementing the infinite variable transmission ratio state is carried out according to learned values—based on an initial setting and learning repeated through driving and stopping thereafter—stored in memory in the controller  66 . However, in the case of the example, the infinitely variable transmission ratio state can be detected on the basis of an output signal from the output shaft rotation speed sensor  68 . Therefore, the stepping motor  13  can also be positioned on the basis of the output signal from the output shaft rotation speed sensor  68 . 
   In any case, during Step  4 , a transmission state of the continuously variable transmission becomes such that the transmission ratio is infinite or nearly infinite; however, correction by the correcting control valve  57  is not yet performed. Under this state, pressurized oil is not supplied to or discharged from the correcting control valve  57  and the differential pressure cylinder  55 , and a spool  60  of the differential pressure cylinder  55  is positioned in its intermediate position. Therefore, a transmission ratio of the toroidal continuously variable transmission  25   a  is not corrected by the correcting control valve  57 . In the state when immediately after a start of engagement of the clutch device  67  and before completion of the engagement, the torque passing through the toroidal continuously variable transmission  25   a  is unstable. However, because the above-mentioned correction of the transmission ratio is not yet performed, the toroidal continuously variable transmission  25   a  cannot be corrected excessively (beyond a required amount). 
   As described above, when the stepping motor  13  is positioned at a position for entering an infinitely variable transmission ratio state simultaneously with the start of engagement of the clutch device  67 , a timer in the controller  66  starts counting (Step  5 ). Thereby, a determination is made as to whether or not a predetermined period of time has elapsed (i.e., a time until the torque becomes stable, which is obtained experimentally in consideration of a temperature during driving, a vehicle-to-vehicle variation, or the like; e.g., a short period of time of one second or less) since the start of engagement of the clutch device  67  (Step  6 ). The clutch device  67  is completely engaged upon elapse of the predetermined time since the start of engagement, whereby the torque passing through the toroidal continuously variable transmission  25   a  becomes stable. Subsequently, correction by the correcting control valve  57  is started (Step  7 ). As a result, as is the case with the constitution of the aforementioned  FIG. 11 , the torque which is transmitted to the output shaft  389  after passing through the toroidal continuously variable transmission  25   a  can be regulated to a desired value. 
   Note that when the input-side rotation speed sensor  64  and the output-side rotation speed sensor  65  are provided as in the case with the example, the passing torque can also be controlled by calculating a rotation speed and a rotation direction of the output shaft  38   a  on the basis of detection signals from the two rotation speed sensors  64 ,  65 . However, in the example, the correction performed by the correcting control valve  57  is arranged so as to start after the torque passing through the toroidal continuously variable transmission  25   a  has become stable. Thereby, the torque under a condition where a non-travel state is switched to a travel state can be controlled more easily and conveniently. In other words, in the example, under a state where the clutch device  67  is disengaged, control for adjusting the transmission ratio of the toroidal continuously variable transmission  25   a  for regulating the torque passing through the toroidal continuously variable transmission  25   a  to a desired value is not implemented by use of detection signals from both the input-side and the output-side rotation speed sensors  64 ,  65 . 
   Next, a control circuit suitable for controlling a continuously variable transmission apparatus of the above mentioned invention will be briefly explained with reference to  FIG. 4 . Note that repeated descriptions are omitted with regard to structures of those portions which have the same structures as those shown in the aforementioned  FIG. 15  and are for controlling a stroke of an actuator  10  to thereby adjust the transmission ratio of the toroidal continuously variable transmission by way of the control valve  12 , the stepping motor  13 , the precess cam  18 , the link arm  19 , and the differential cylinder  55 . 
   In a hydraulic circuit shown in  FIG. 4 , pressure of the pressurized oil suctioned from oil sumps  81  and then discharged from oil pumps  69   a ,  69   b  can be adjusted to a predetermined pressure by means of pressure regulating valves  82   a ,  82   b . The oil pumps  69   a ,  69   b  correspond to the aforementioned oil pump  69  in  FIG. 2 . Further, with regard to the two pressure regulating valves  82   a ,  82   b , an adjustment pressure—applied to the pressure regulating valve  82   a  for adjusting an pressure of oil sent to an oil-pressure-manual-switch valve  79  side to be described later—can be adjusted by means of opening and closing the line-pressure-control solenoid valve  73 . Further, the pressurized oil, whose pressure is adjusted by the two pressure regulating valves  82   a ,  82   b , can be sent to the actuator  10  via a control valve  12 . The pressurized oil is also sent to the correcting control valve  57  for adjusting a stroke of the differential pressure cylinder  55  by means of opening and closing the solenoid valves  58   a ,  58   b.    
   Moreover, the pressurized oil is arranged so as to be sent to the hydraulic loader  23   a . The pressurized oil can also be sent to an oil chamber of the low-speed clutch  40   a  ( 40 ) or the high-speed clutch  41   a  ( 41 ) via the oil-pressure-manual-switch valve  79 , and the high-speed switch valve  71  or the low-speed switch valve  72 . Of the valves  79 ,  71 , and  72 , the oil-pressure-manual-switch valve  79  is switched by a control lever (i.e., shift lever), which is provided at a driver&#39;s seat and controlled by a driver, for selecting a parking range (P), a reverse (i.e., backward) range (R), a neutral range (N), a drive (generally forward) range (D), or a forward-with-high-driving-force range (L).  FIG. 4  shows the respective switching states of the oil-pressure-manual-switch valve  79  under the states where the respective ranges are selected. Note that structures and functions of the respective valves including the oil-pressure-manual-switch valve  79  are represented in accordance with a general form of engineering drawing of hydraulic equipment. 
   Communication states of the high-speed and low-speed switch valves  71 ,  72  are respectively switched upon supply and discharge of pressurized oil on the basis of switching of a shift valve  83  switched by the shift solenoid valve  74 . When one of the valves  71  (or  72 ) sends pressurized oil to an oil chamber of the high-speed clutch  41   a  (or an oil chamber of the low-speed clutch  40   a ), the other valve  72  (or  71 ) discharges pressurized oil from the oil pressure chamber of the low-speed clutch  40   a  (or the oil chamber of the high-speed clutch  41   a ). 
   A controller which is provided with a hydraulic circuit constituted as described above and which is incorporated in a continuously variable transmission apparatus constituted as shown in the aforementioned  FIGS. 1 and 2  has the following functions (a) through (f): 
   (a) a function of adjusting a transmission ratio of the toroidal continuously variable transmission  25   a  so as to change a relative displacement speed of a plurality of gears constituting the planetary-gear-type transmission  26   b , thereby changing the rotation condition of the output shaft  38   a  between the forward-rotating condition and the reverse-rotating condition through the non-rotating condition while the input shaft  1  is kept rotating in a single direction in a low-speed mode by the engine  62  functioning as a driving source; that is, during a driving state where the low-speed clutch  40   a  is engaged and the high-speed clutch  41   a  is disengaged 
   (function (a) is the same as that of a continuously variable transmission having a conventionally known constitution shown in aforementioned  FIG. 10 , or the structure shown in  FIG. 11 ); 
   (b) a function of changing a transmission ratio between the input shaft  1  and the output shaft  38   a  to thereby change the transmission ratio of the toroidal continuously variable transmission  25   a  in a high-speed mode; that is, during a driving state where the low-speed clutch  40   a  is disengaged and the high-speed clutch  41   a  is engaged 
   (function (b) is also the same as that of a continuously variable transmission having a conventionally known constitution shown in aforementioned  FIG. 10 , or the structure shown in  FIG. 11 .); 
   (c) a function of adjusting a torque passing through the toroidal continuously variable transmission  25   a  to thereby change the transmission ratio of the toroidal continuously variable transmission  25   a  in a low-speed mode; that is, during a driving state where the low-speed clutch  40   a  is engaged and the high-speed clutch  41   a  is disengaged {function (2) of a controller}; 
   (d) a function of disengaging both the low-speed clutch  40   a  and the high-speed clutch  41   a  under a non-travel state; that is, a state where a parking range or a neutral range is selected by means of the control lever; 
   (e) a function of reducing a pressing force generated by the loader  23   a  during a vehicle stop or during very low speed driving so that the pressing force becomes smaller than that generated during normal driving; and 
   (f) a function, during a vehicle stop or very low speed driving and when braking means used for stopping the vehicle is applied, of reducing a torque passing through the toroidal continuously variable transmission  25   a  so that the torque becomes smaller than that under a condition where the braking means is not applied thereon. 
   [Second Embodiment] 
   Next,  FIG. 5  shows a flow chart for describing operation of a second embodiment of the invention. In the second embodiment, correction by a correcting control valve  57  (see  FIG. 4 ) is started after an oil pressure differential between a pair of oil pressure chambers  24   a ,  24   b  provided in an actuator  10  has reached a predetermined value. As described above, an oil pressure differential between the pair of oil pressure chambers  24   a ,  24   b  is proportional to a torque passing through the toroidal continuously variable transmission  25   a  (see  FIGS. 1 and 2 ). Accordingly, by observing the oil pressure differential, engagement of a clutch device  67  ( FIG. 2 ) and stabilization of the torque can be detected. Accordingly, the correction is arranged so as to start when the oil pressure differential indicates that the torque has become stable. After the correction is started, the torque which is transmitted to the output shaft  38   a  after passing through the toroidal continuously variable transmission  25  (see  FIGS. 1 and 2 ) is regulated to a desired value, as in the case of the aforementioned structure shown in  FIG. 11  and the aforementioned first embodiment. 
   [Third Embodiment] 
   Next,  FIG. 6  shows a flow chart for describing operation of a third embodiment of the invention. In the third embodiment, a load-detecting device for detecting a load on the driving source engine  62  ( FIG. 2 ) is provided. The engine controller  80  ( FIG. 2 ) can be utilized as the load-detecting unit. That is, an increase in the load on the engine can be detected by utilizing a signal for directing that the fuel amount to be supplied to the engine should be increased, which is output from the engine controller  80  when a load on the engine  62  is increased. Then, on the basis of the increase in the load on the engine  62 , engagement of the clutch device  67  ( FIG. 2 ) and stabilization of the torque passing through the toroidal continuously variable transmission  25   a  (see  FIGS. 1 and 2 ) can be detected. Therefore, the correction is arranged so as to start when the increase in the load of the engine  62  indicates that the torque has become stable. After the correction is started, the torque which is transmitted to the output shaft  38   a  after passing through the toroidal continuously variable transmission  25  (see  FIGS. 1 and 2 ) is regulated to a desired value, as in the case of the aforementioned structure shown in  FIG. 11  and the aforementioned first embodiment.