JP4066920B2 - Testing equipment for toroidal type continuously variable transmissions - Google Patents

Description

    The present invention tests the performance of a transmission ratio control unit for controlling a transmission ratio of a toroidal continuously variable transmission that is combined with a gear-type differential mechanism such as a planetary gear transmission to form a continuously variable transmission. The present invention relates to a test apparatus used for the purpose.

  The use of a toroidal type continuously variable transmission as shown in FIGS. 5 to 7 as an automatic transmission for automobiles has been studied and implemented in part. This toroidal type continuously variable transmission is called a double cavity type, and supports input-side disks 2 and 2 around both ends of the input shaft 1 via ball splines 3 and 3. Therefore, both the input side disks 2 and 2 are supported concentrically and freely in a synchronized manner. An output gear 4 is supported around the intermediate portion of the input shaft 1 so as to be freely rotatable relative to the input shaft 1. The output side disks 5 and 5 are respectively spline-engaged with both ends of a cylindrical portion provided at the center of the output gear 4. Accordingly, both the output side disks 5 and 5 rotate in synchronism with the output gear 4.

  A plurality (usually 2 to 3) of power rollers 6 and 6 are sandwiched between the input disks 2 and 2 and the output disks 5 and 5, respectively. Each of these power rollers 6 and 6 is rotatably supported by inner surfaces of trunnions 7 and 7 which are support members via support shafts 8 and 8 and a plurality of rolling bearings. The trunnions 7 and 7 are pivotal shafts 9 and 9 provided concentrically with each other for each trunnion 7 and 7 at both ends in the length direction (the vertical direction in FIGS. 5 and 7 and the front and back direction in FIG. 6). Oscillating and displacing around the center. The operation of inclining the trunnions 7 and 7 is performed by displacing the trunnions 7 and 7 in the axial direction of the pivots 9 and 9 by the hydraulic actuators 10 and 10. The inclination angle of 7 is synchronized with each other hydraulically and mechanically.

  That is, when changing the inclination angle of the trunnions 7 and 7 in order to change the transmission ratio between the input shaft 1 and the output gear 4, the trunnions 7 and 7 are moved by the actuators 10 and 10, respectively. In the opposite directions, for example, the power roller 6 on the right side of FIG. 7 is displaced to the lower side in the figure, and the power roller 6 on the left side in the figure is displaced to the upper side in the figure by the same distance. As a result, the direction of the tangential force acting on the contact portion between the peripheral surface of each of the power rollers 6 and 6 and the inner surface of each of the input side disks 2 and 2 and the output side disks 5 and 5 changes. (Side slip occurs at the contact portion). Then, the trunnions 7 and 7 swing (tilt) in opposite directions around the pivots 9 and 9 pivotally supported by the support plates 11 and 11 as the direction of the force changes. As a result, the contact position between the peripheral surface of each of the power rollers 6 and 6 and the inner surface of each of the input and output disks 2 and 5 changes, and rotation between the input shaft 1 and the output gear 4 occurs. The gear ratio changes.

  Regardless of the number of these actuators 10, 10, the supply / discharge state of the pressure oil to each of the actuators 10, 10 is performed by one control valve 12, and any one trunnion 7 is moved. I'm trying to provide feedback. This control valve 12 is fitted in a sleeve 14 that is displaced in the axial direction (left-right direction in FIG. 7, the front-back direction in FIG. 5) by a stepping motor 13 and is axially displaceable on the inner diameter side of the sleeve 14. And a spool 15. A precess cam 18 is attached to the end of the rod 17 attached to any one of the trunnions 7 among the rods 17 and 17 connecting the trunnions 7 and 7 and the pistons 16 and 16 of the actuators 10 and 10. A feedback mechanism that is fixed and transmits the movement of the rod 17, that is, the combined value of the displacement amount in the axial direction and the displacement amount in the rotation direction, to the spool 15 via the recess cam 18 and the link arm 19. Is configured. Further, a synchronizing cable 20 is spanned between the trunnions 7 and 7 so that the inclination angles of the trunnions 7 and 7 can be mechanically synchronized even when the hydraulic system fails.

  When switching the speed change state, the stepping motor 13 displaces the sleeve 14 to a predetermined position corresponding to the speed ratio to be obtained, and opens the flow path of the control valve 12 in a predetermined direction. As a result, pressure oil is sent to the actuators 10 and 10 in a predetermined direction, and the actuators 10 and 10 displace the trunnions 7 and 7 in a predetermined direction. That is, the trunnions 7 and 7 swing around the pivots 9 and 9 while being displaced in the axial direction of the pivots 9 and 9 as the pressure oil is fed. Then, the movement (axial direction and swing displacement) of any one of the trunnions 7 is transmitted to the spool 15 via a recess cam 18 and a link arm 19 fixed to the end of the rod 17, and this spool 15 is displaced in the axial direction. As a result, in the state where the trunnion 7 is displaced by a predetermined amount, the flow path of the control valve 12 is closed, and the supply and discharge of the pressure oil to the actuators 10 and 10 are stopped.

  At this time, the movement of the control valve 12 based on the displacement of the cam surface 21 of the trunnion 7 and the recess cam 18 is as follows. First, when the trunnion 7 is displaced in the axial direction as the flow path of the control valve 12 is opened, as described above, the peripheral surface of the power roller 6 and the inner surfaces of the input side disk 2 and the output side disk 5. The trunnion 7 starts swinging displacement about the pivots 9 and 9 due to the side slip generated at the contact part. Further, the displacement of the cam surface 21 is transmitted to the spool 15 through the link arm 19 in accordance with the axial displacement of the trunnion 7, and the spool 15 is displaced in the axial direction to switch the control valve 12. Change state. Specifically, the control valve 12 is switched in a direction to return the trunnion 7 to the neutral position by the actuator 10.

  Therefore, immediately after the trunnion 7 is displaced in the axial direction, the trunnion 7 starts to be displaced in the opposite direction toward the neutral position. However, the trunnion 7 continues to swing around the pivots 9 and 9 as long as there is a displacement from the neutral position. As a result, the displacement in the circumferential direction of the cam surface 21 of the recess cam 18 is transmitted to the spool 15 via the link arm 19, and the spool 15 is displaced in the axial direction. Then, in a state where the inclination angle of the trunnion 7 has reached a predetermined angle corresponding to the speed ratio to be obtained, the trunnion 7 returns to the neutral position, and at the same time, the control valve 12 is closed to the actuator 10. The supply and discharge of pressure oil is stopped. As a result, the inclination angle of the trunnion 7 becomes an angle commensurate with the amount of displacement of the sleeve 14 in the axial direction by the stepping motor 13.

  During operation of the toroidal type continuously variable transmission as described above, one input side disk 2 (left side in FIGS. 5 and 7) is connected to a loading cam type as shown by a drive shaft 22 connected to a power source such as an engine. It is rotationally driven via the pressing device 23. As a result, the pair of input side disks 2 and 2 supported at both ends of the input shaft 1 rotate synchronously while being pressed toward each other. Then, the rotation is transmitted to the output side disks 5 and 5 through the power rollers 6 and 6 and is taken out from the output gear 4.

  As described above, when power is transmitted from the input disks 2 and 2 to the output disks 5 and 5, the trunnions 7 and 7 have power troughs 6 and 7 supported on inner surfaces thereof. Along with the friction at the rolling contact portion (traction portion) between the peripheral surface of the disk 6 and the inner surface of each of the disks 2 and 5, axial forces of the pivots 9 and 9 provided at both ends are applied. This force is so-called 2Ft, and the magnitude of the force is from the input disks 2 and 2 to the output disks 5 and 5 (or from the output disks 5 and 5 to the input disks 2 and 2). Is proportional to the torque transmitted to. Such a force 2Ft is supported by the actuators 10 and 10. Therefore, when the toroidal continuously variable transmission is operated, the pressure difference between the pair of hydraulic chambers 24a and 24b existing on both sides of the pistons 16 and 16 constituting the actuators 10 and 10 is the force 2Ft. Proportional to size.

  When the rotational speed of the input shaft 1 and the output gear 4 is changed, and when the deceleration is first performed between the input shaft 1 and the output gear 4, the trunnions 7 and 7 are moved by the actuators 10 and 10, respectively. The trunnions 9 and 9 are moved in the axial direction, and the trunnions 7 and 7 are swung to the positions shown in FIG. Then, as shown in FIG. 6, the peripheral surfaces of the power rollers 6 and 6 are arranged near the center of the inner surface of the input disks 2 and 2 and the outer periphery of the inner surfaces of the output disks 5 and 5, respectively. It abuts on each side part. On the contrary, when the speed is increased, the trunnions 7 and 7 are swung in the direction opposite to that shown in FIG. 6, and the peripheral surfaces of the upper power rollers 6 and 6 are opposite to the state shown in FIG. In addition, the trunnions 7 and 7 are inclined so as to abut the outer peripheral portions of the inner side surfaces of the input side disks 2 and 2 and the central portion of the inner side surfaces of the output side disks 5 and 5, respectively. Let If the inclination angles of the trunnions 7 and 7 are set in the middle, an intermediate speed ratio (speed ratio) can be obtained between the input shaft 1 and the output gear 4.

  Furthermore, when the toroidal continuously variable transmission configured and operated as described above is incorporated into an actual continuously variable transmission for an automobile, a continuously variable transmission is configured in combination with a gear-type differential unit such as a planetary gear mechanism. It has been proposed in the past. For example, Patent Document 1 discloses a continuously variable transmission called so-called geared neutral, in which the rotation state of the output shaft can be switched between forward rotation and reverse rotation with the input shaft rotated in one direction with the stop state interposed therebetween. Are listed. FIG. 8 shows a continuously variable transmission described in Patent Document 1 as described above. This continuously variable transmission comprises a combination of a toroidal type continuously variable transmission 25 and a planetary gear type transmission 26. Of these, the toroidal-type continuously variable transmission 25 includes an input shaft 1, a pair of input-side disks 2, 2, an output-side disk 5 a, and a plurality of power rollers 6, 6. In the illustrated example, the output side disk 5a has a structure such that the outer surfaces of a pair of output side disks are brought into contact with each other and integrated.

  The planetary gear type transmission 26 includes a carrier 27 coupled and fixed to the input shaft 1 and the input side disk 2 on one side (right side in FIG. 8). A first transmission shaft 29 in which planetary gear elements 28a and 28b are fixed to both ends of the carrier 27 in the radial direction is supported rotatably. Further, on the opposite side of the input shaft 1 across the carrier 27, a second transmission shaft 31 having sun gears 30a and 30b fixed to both ends thereof is supported concentrically with the input shaft 1 so as to be rotatable. is doing. Then, the planetary gear elements 28a and 28b and the distal end portion (the right end portion in FIG. 8) of the hollow rotary shaft 32 having the base end portion (the left end portion in FIG. 8) coupled to the output side disk 5a are fixed. The sun gear 33 or the sun gear 30a fixed to one end (the left end in FIG. 8) of the second transmission shaft 31 is meshed with each other. One planetary gear element 28a (left side in FIG. 8) is meshed with a ring gear 35 rotatably provided around the carrier 27 via another planetary gear element 34.

  On the other hand, the planetary gear elements 37a and 37b are rotatably supported by the second carrier 36 provided around the sun gear 30b fixed to the other end portion (the right end portion in FIG. 8) of the second transmission shaft 31. is doing. The second carrier 36 is fixed to the base end portion (left end portion in FIG. 8) of the output shaft 38 disposed concentrically with the input shaft 1 and the second transmission shaft 31. The planetary gear elements 37a and 37b mesh with each other, and one planetary gear element 37a is rotatable around the sun gear 30b and the other planetary gear element 37b is rotatable around the second carrier 36. The second ring gear 39 provided is meshed with each other. The ring gear 35 and the second carrier 36 can be freely engaged and disengaged by a low speed clutch 40, and the second ring gear 39 and a fixed part such as a housing are engaged by a high speed clutch 41. It is considered to be removable.

  In the case of the continuously variable transmission shown in FIG. 8 as described above, in the so-called low speed mode state in which the low speed clutch 40 is connected and the high speed clutch 41 is disconnected, the power of the input shaft 1 is This is transmitted to the output shaft 38 via the ring gear 35. By changing the gear ratio of the toroidal continuously variable transmission 25, the gear ratio of the continuously variable transmission, that is, the gear ratio between the input shaft 1 and the output shaft 38 changes. In such a low speed mode state, the speed ratio of the continuously variable transmission as a whole changes to infinity. That is, by adjusting the gear ratio of the toroidal-type continuously variable transmission 25, the rotation state of the output shaft 38 is rotated forward with the stop state between the input shaft 1 and the input shaft 1 rotated in one direction. , Reverse conversion is possible.

  Incidentally, during acceleration or constant speed running in such a low speed mode state, the torque (passing torque) passing through the toroidal continuously variable transmission 25 is transferred from the input shaft 1 to the carrier 27 and the first transmission shaft 29. And the sun gear 33 and the hollow rotating shaft 32 are added to the output side disk 5a, and further, the output side disk 5a is added to the input side disks 2 and 2 via the power rollers 6 and 6, respectively. That is, the torque passing through the toroidal type continuously variable transmission 25 during acceleration or constant speed circulation circulates in a direction in which the input disks 2 and 2 receive torque from the power rollers 6 and 6.

On the other hand, in the so-called high speed mode state in which the low speed clutch 40 is disconnected and the high speed clutch 41 is connected, the power of the input shaft 1 causes the first and second transmission shafts 29 and 31 to be connected. Via the output shaft 38. Then, by changing the gear ratio of the toroidal-type continuously variable transmission 25, the gear ratio of the entire continuously variable transmission changes. In this case, the greater the gear ratio of the toroidal type continuously variable transmission 25, the greater the gear ratio of the continuously variable transmission.
Incidentally, the torque passing through the toroidal type continuously variable transmission 25 during acceleration or constant speed traveling in such a high speed mode state is that each input disk 2, 2 applies torque to each power roller 6, 6. Join the direction.

  For example, in the case of a continuously variable transmission having a structure as shown in FIG. 8 and capable of realizing a so-called infinite transmission ratio in which the output shaft 38 is stopped while the input shaft 1 is rotated, the output shaft 38 is Maintaining the torque applied to the toroidal continuously variable transmission 25 at an appropriate value in a state where the gear ratio is extremely large, including the stopped state, ensures durability of the toroidal continuously variable transmission 25. This is important in terms of ensuring the ease of operation. As is clear from the relationship of “rotational driving force = rotational speed × torque”, in the state where the gear ratio is extremely large and the output shaft 38 is stopped or rotated at an extremely low speed while the input shaft 1 is rotating. The torque (passing torque) that passes through the toroidal-type continuously variable transmission 25 becomes larger than the torque applied to the input shaft 1. Therefore, in order to ensure the durability of the toroidal type continuously variable transmission 25 without increasing the size of the toroidal type continuously variable transmission 25, it is strictly necessary to keep the torque at an appropriate value as described above. It becomes necessary to perform control. Specifically, in order to stop the output shaft 38 while reducing the torque input to the input shaft 1 as much as possible, control including the drive source is required.

  Further, when the gear ratio is extremely large, even when the gear ratio of the toroidal continuously variable transmission 25 slightly changes, the torque applied to the output shaft 38 changes greatly. For this reason, if the gear ratio adjustment of the toroidal-type continuously variable transmission 25 is not strictly performed, there is a possibility that the driver feels uncomfortable or that the driving operation becomes difficult. For example, in the case of an automatic transmission for an automobile, when the vehicle is stopped, the driver keeps stepping on the brake and maintains the stopped state. In such a case, the gear ratio adjustment of the toroidal-type continuously variable transmission 25 is not performed strictly, and if a large torque is applied to the output shaft 38, the force required to depress the brake pedal when the vehicle is stopped increases. Increases driver fatigue. On the contrary, if the gear ratio of the toroidal-type continuously variable transmission 25 is not strictly adjusted at the time of starting, and if the torque applied to the output shaft 38 is too small, smooth starting cannot be performed or when starting on an uphill The vehicle may move backwards. Therefore, at the time of stopping or traveling at an extremely low speed, it is necessary to strictly control the transmission ratio of the toroidal continuously variable transmission 25 in addition to controlling the torque transmitted from the drive source to the input shaft 1.

In consideration of such points, Patent Document 2 discloses a torque (passing torque) that passes through a toroidal continuously variable transmission by directly controlling a pressure difference of a hydraulic actuator for displacing a trunnion. The structure which regulates is described.
However, in the case of the structure as described in Patent Document 2, since the control is performed only by the pressure difference, it is difficult to stop the trunnion posture at the moment when the passing torque matches the target value. Specifically, since the amount of displacement of the trunnion for torque control increases, the so-called overshoot (further, the trunnion continues to displace without stopping when the passing torque matches the target value. Hunting associated with this is likely to occur, and the control of the passing torque is not stable.

  In particular, like the general half-toroidal continuously variable transmission shown in FIGS. 5 to 7, the directions of the pivots 9 and 9 provided at both ends of the trunnions 7 and 7, and the input-side and output-side discs 2. In the case of the toroidal-type continuously variable transmission 25 having a so-called cast angle in which the directions of the central axes of 5 are perpendicular to each other, the overshoot is likely to occur. On the other hand, in the case of a structure having a cast angle like a general full toroidal type continuously variable transmission, a force in a direction to converge the overshoot acts, and therefore, it is described in Patent Document 2 described above. Even with such a structure, it is considered that sufficient torque control can be performed.

In view of these circumstances, Japanese Patent Application No. 2003-56681 discloses a continuously variable transmission that incorporates a toroidal continuously variable transmission that does not have a cast angle, such as a general half-toroidal continuously variable transmission. In addition, a method and an apparatus capable of strictly controlling the torque (passing torque) passing through the toroidal continuously variable transmission are disclosed.
FIG. 9 shows an example of the structure of a continuously variable transmission that is the object of the control method and apparatus of the prior invention. The continuously variable transmission shown in FIG. 9 has the same function as that of the conventionally known continuously variable transmission shown in FIG. 8, but the structure of the planetary gear type transmission 26a portion is the same. By devising, the assemblability of this planetary gear type transmission 26a is improved.

  First and second planetary gears 42 and 43, each of which is a double pinion type, are supported on both side surfaces of a carrier 27a that rotates together with the input shaft 1 and the pair of input side disks 2 and 2. That is, each of the first and second planetary gears 42 and 43 includes a pair of planetary gear elements 44a, 44b, 45a, and 45b. These planetary gear elements 44a, 44b, 45a, 45b are engaged with each other, and the planetary gear elements 44a, 45a on the inner diameter side are coupled to the output side disk 5a at the base end portion (left end portion in FIG. 9). The first and second sun gears 47 and 48 fixed to the tip end portion (right end portion in FIG. 9) of the hollow rotating shaft 32a and one end portion (left end portion in FIG. 9) of the transmission shaft 46 are connected to the outer diameter side. Planetary gear elements 44b and 45b are meshed with the ring gear 49, respectively.

  On the other hand, the planetary gear elements 51a and 51b are rotatably supported by the second carrier 36a provided around the third sun gear 50 fixed to the other end portion (the right end portion in FIG. 9) of the transmission shaft 46. is doing. The second carrier 36a is fixed to the base end portion (left end portion in FIG. 9) of the output shaft 38a arranged concentrically with the input shaft 1. The planetary gear elements 51a and 51b mesh with each other, and the planetary gear element 51a on the inner diameter side is connected to the third sun gear 50, and the planetary gear element 51b on the outer diameter side is connected to the second carrier 36a. Are respectively meshed with second ring gears 39a that are rotatably provided around the gears. Further, the ring gear 49 and the second carrier 36a can be freely engaged and disengaged by a low speed clutch 40a, and the second ring gear 39a and a fixed part such as a housing are engaged by a high speed clutch 41a. It is considered to be removable.

In the improved continuously variable transmission configured as described above, the power of the input shaft 1 is transmitted via the ring gear 49 in a state where the low speed clutch 40 a is connected and the high speed clutch 41 a is disconnected. Is transmitted to the output shaft 38a. Then, by changing the gear ratio of the toroidal-type continuously variable transmission 25, the speed ratio e _CVT of the continuously variable transmission, that is, the speed ratio between the input shaft 1 and the output shaft 38a changes. The relationship between the speed ratio e _CVU of the toroidal-type continuously variable transmission 25 and the speed ratio e _CVT of the continuously variable transmission as a whole is that the number of teeth m ₄₉ of the ring gear ₄₉ and the first sun gear 47 are When the ratio with the number of teeth m ₄₇ is i ₁ (= m ₄₉ / m ₄₇ ), it is expressed by the following equation (1).
e _CVT = (e _CVU + i ₁ -1) / i ₁ --- (1)
For example, when the ratio i ₁ between the number of teeth is 2, the relationship between the speed ratios e _CVU and e _CVT changes as shown by a line segment α in FIG.

In contrast, when the low speed clutch 40a is disconnected and the high speed clutch 41a is connected, the power of the input shaft 1 is the first planetary gear 42, the ring gear 49, and the second gear. The planetary gear 43, the transmission shaft 46, the planetary gear elements 51a and 51b, and the second carrier 36a are transmitted to the output shaft 38a. Then, by changing the speed ratio e _CVU of the toroidal type continuously variable transmission 25, the speed ratio e _CVT of the continuously variable transmission as a whole changes. At this time, the relationship between the speed ratio e _CVU of the toroidal-type continuously variable transmission 25 and the speed ratio e _CVT of the continuously variable transmission as a whole is expressed by the following equation (2). Incidentally, the ratio in (2), i ₁ and the number of teeth m ₄₇ teeth number m ₄₉ and the first sun gear 47 of the ring gear _{49 (m 49 / m 47)} , i 2 is the ring the ratio between the teeth number m ₄₉ teeth number m ₄₈ of the second sun gear 48 of the gear 49 _{(m 49 / m 48),} i 3 is the number of teeth m ₃₉ of the second ring gear 39a first the ratio between the teeth number m ₅₀ of third sun gear 50 (m ₃₉ / m _50), represent respectively.
e _CVT = {1 / (1-i ₃ )} · {1+ (i ₂ / i ₁ ) (e _CVU −1)} --- (2)
Of the above ratios, when i ₁ is 2, i ₂ is 2.2, and i ₃ is 2.8, the relationship between the speed ratios e _CVU and e _CVT is shown in FIG. It changes as indicated by β.

  In the case of the continuously variable transmission constructed and operated as described above, as is apparent from the line segment α in FIG. 10, the output shaft 38a is stopped while the input shaft 1 is rotated, so-called infinite gear ratio. The state of can be created. However, when the output shaft 38a is stopped or rotated at an extremely low speed while the input shaft 1 is rotated as described above, it passes through the toroidal type continuously variable transmission 25 as described above. Torque (passing torque) becomes larger than the torque applied to the input shaft 1 from the engine as the drive source. For this reason, it is necessary to properly regulate the torque input to the input shaft 1 from the drive source so that the passing torque does not become excessive (or excessive) when the vehicle is stopped or operated at a low speed. .

  Further, during the slow speed operation, the output shaft 38a is in a state close to being stopped, that is, the gear ratio of the continuously variable transmission is very large, and the rotational speed of the output shaft 38a is higher than the rotational speed of the input shaft 1. In a state where the speed is significantly slow, the torque applied to the output shaft 38a varies greatly due to a slight variation in the gear ratio of the continuously variable transmission. For this reason, in order to ensure a smooth driving operation, it is also necessary to properly regulate the torque input to the input shaft 1 from the drive source.

  Incidentally, during acceleration or constant speed running in such a low speed mode state, the passing torque is transmitted from the input shaft 1 to the carrier 27a, the first planetary gear 42 and the first gear as in the conventional structure shown in FIG. It is added to the output side disk 5a through one sun gear 47 and the hollow rotary shaft 32a, and further from the output side disk 5a to each input side disk 2, 2 through each power roller 6, 6 (see FIG. 8). Join. That is, during acceleration or constant speed travel, the passing torque circulates in a direction in which the input disks 2 and 2 receive torque from the power rollers 6 and 6.

  For this reason, in the case of the speed ratio control method and apparatus according to the previous invention, the torque input to the input shaft 1 from the drive source is appropriately regulated as shown in FIG. First, the rotational speed of the engine that is the drive source is roughly controlled. That is, the rotational speed of the engine is restricted to a point a within the range w of FIG. At the same time, the transmission gear ratio of the toroidal continuously variable transmission 25 required to make the rotation speed of the input shaft 1 of the continuously variable transmission match the controlled engine rotation speed is set. This setting operation is performed based on the above-described equation (1). That is, it is necessary to strictly regulate the torque transmitted from the engine to the input shaft 1 by the method of the previous invention. The low speed clutch 40a is connected and the high speed clutch 41a is disconnected, so-called low speed. In mode. Therefore, the gear ratio of the toroidal continuously variable transmission 25 is set according to the above equation (1) so that the rotational speed of the input shaft 1 is a value corresponding to the required rotational speed of the output shaft 38a.

Further, a pair of hydraulic chambers 24a and 24b (see FIG. 5) constituting the hydraulic actuators 10 and 10 for displacing the trunnions 7 and 7 incorporated in the toroidal type continuously variable transmission 25 in the axial direction of the pivots 9 and 9. 7 and FIG. 13 described later) is measured by a hydraulic sensor (not shown). In this hydraulic pressure measurement operation, the rotational speed of the engine is controlled roughly (however, the rotational speed is kept constant). Correspondingly, as described above, the toroidal type continuously variable transmission according to the equation (1). This is performed in a state where a gear ratio of 25 is set. Then, the torque (passing torque) T _CVU passing through the toroidal type continuously variable transmission 25 is calculated from the pressure difference obtained based on the measurement work.

That is, the pressure difference is proportional to the torque T _CVU passing through the toroidal type continuously variable transmission 25 as long as the transmission ratio of the toroidal type continuously variable transmission 25 is constant. T _CVU can be calculated. The reason for this is that, as described above, the torque transmitted by the actuators 10 and 10 from the input side disks 2 and 2 to the output side disk 5a (or from the output side disk 5a to the input side disks 2 and 2) ( This is because a force of 2 Ft having a magnitude proportional to the torque T _CVU passing through the toroidal-type continuously variable transmission 25) is supported.

On the other hand, the torque T _CVU can be obtained by the following equation (3).
T _CVU = e _CVU · T _IN / {e _CVU + (i ₁ -1) η _CVU } --- (3)
In this equation (3), e _CVU is the speed ratio of the toroidal type continuously variable transmission 25, T _IN is the torque input to the input shaft 1 from the engine, and i ₁ is related to the first planetary gear 42. gear ratio of the planetary gear transmission (the ratio between the teeth number m ₄₇ of teeth number m ₄₉ and the first sun gear 47 of the ring gear 49), eta _CVU is the efficiency of the toroidal-type continuously variable transmission 25, Represents each.

Therefore, calculated from the pressure difference, the torque T _CVU1 actually passing through the toroidal type continuously variable transmission 25, was determined from equation (3) based on the passing torque T _CVU2 a target, the actual A deviation ΔT (= T _CVU1 −T _CVU2 ) between the passing torque T _CVU1 and the target value T _CVU2 is _obtained . Then, the speed ratio of the toroidal continuously variable transmission 25 is adjusted in a direction to eliminate the deviation ΔT (assuming ΔT = 0). Since the torque deviation ΔT and the pressure difference deviation are in a proportional relationship, the gear ratio adjustment operation can be performed by either a torque deviation or a pressure difference deviation. That is, the transmission ratio control based on the torque deviation and the transmission ratio control based on the pressure difference deviation are technically the same.

For example, as shown in FIG. 11, the engine drives the input shaft 1 in a region where the torque T _CVU1 (measured value) that actually passes through the toroidal-type continuously variable transmission 25 is regulated to a target value T _CVU2. Let us consider a case where the torque T _IN changes in a direction that rapidly decreases as the rotational speed of the input shaft 1 increases. Such engine characteristics can be easily obtained even in a low-speed rotation region if the engine is electronically controlled. In the case of such engine characteristics, the measured torque value T _CVU1 of the above-mentioned is similar to the target value T _CVU2 . In order to reduce the torque T _IN for driving the input shaft 1, the speed ratio of the continuously variable transmission as a whole is displaced to the deceleration side in order to reduce the torque T _IN for driving the input shaft 1. . For this purpose, the speed ratio of the toroidal type continuously variable transmission 25 is shifted to the speed increasing side.

  However, when the brake pedal is depressed (rotation speed of the output shaft = 0), the slip generated inside the toroidal-type continuously variable transmission 25, that is, the inner surfaces of the disks 2 and 5a on the input side and output side, respectively. And the speed ratio of the toroidal continuously variable transmission 25 within a range that can be absorbed by a slip (creep) generated at a contact portion (traction portion) between the power roller 6 and the peripheral surface of each power roller 6, 6 (see FIGS. 6 to 8). Control. Therefore, the allowable range in which the speed ratio can be adjusted is limited to a range in which an excessive force is not applied to the contact portion, and is limited as compared with the case of low speed traveling.

For example, in FIG. 11, when the target value T _CVU2 is present at the point a and the measured value T _CVU1 is present at the point b in FIG. 6 has a deviation in the direction of receiving torque. Therefore, the speed ratio e _CVU of the toroidal continuously variable transmission 25 is changed to the speed increasing side, and the speed ratio e _CVT of the continuously variable transmission (T / M) as a whole is changed to the speed reducing side. In accordance with this, the rotational speed of the engine is increased and the torque is decreased. On the contrary, when the measured value T _CVU1 is present at the point c in the figure, each of the input side disks 2 and 2 has a deviation in a direction in which torque is applied to the power rollers 6 and 6. In this case, contrary to the case described above, the speed ratio e _CVU of the toroidal continuously variable transmission 25 is changed to the deceleration side, and the speed ratio e _{CVT of the} continuously variable transmission (T / M) as a whole is changed. Change to the speed increasing side. In accordance with this, the rotational speed of the engine is reduced to increase the torque.

Thereafter, the above-described operation is repeated until the torque T _CVU1 that actually passes through the toroidal-type continuously variable transmission 25 determined from the pressure difference matches the target value. That is, if the torque T _CVU1 passing through the toroidal continuously variable transmission 25 cannot be made equal to the target value T _CVU2 by only one shift control of the toroidal continuously variable transmission 25, the above-described operation is performed. Repeat. As a result, the torque T _IN for rotationally driving the input shaft 1 by the engine can be brought close to a value for _setting the torque T _CVU passing through the toroidal-type continuously variable transmission 25 to the target value T _CVU2 . Such an operation is performed automatically and in a short time in response to a command from a microcomputer incorporated in the controller of the continuously variable transmission.

FIG. 12 shows the ratio (the left vertical axis) of the torque T _CVU passing through the toroidal type continuously variable transmission 25 and the torque T _IN when the engine rotates the input shaft 1, and the entire continuously variable transmission. The relationship between the speed ratio e _CVT (horizontal axis) and the speed ratio e _CVU (right vertical axis) of the toroidal-type continuously variable transmission 25 is shown. The solid line a is as above passing torque T _CVU and the ratio between the drive torque T _IN, the relationship between the speed ratio e _CVT of the total of the continuously variable transmission device, the dashed line b is the both the speed ratio e _CVT, the relationship between e _CVU , Respectively. The prior invention represents the torque T _CVU1 actually passing through the toroidal continuously variable transmission 25 as a point on the solid line a in a state where the speed ratio e _CVT of the continuously variable transmission as a whole is regulated to a predetermined value. The speed ratio e _CVU of the toroidal-type continuously variable transmission 25 is regulated in order to regulate the target value (T _CVU2 ).

In the case of the prior invention, the control for restricting the torque T _CVU1 actually passing through the toroidal-type continuously variable transmission 25 to the point on the solid line a that is the target value T _CVU2 is divided into two stages. That is, the engine speed is roughly controlled, that is, the target value T _CVU2 is controlled to be considered to be _able to obtain the target value T _CVU2 , and then the gear ratio control of the toroidal continuously variable transmission 25 is performed in accordance with this speed. To do. For this reason, the toroidal-type continuously variable transmission 25 is not actually generated without causing an overshoot (and hunting associated therewith) as in the conventional method, or even if it occurs, to a practically low level. The passing torque T _CVU1 can be regulated to the target value T _CVU2 .

  As described above, the driving force (torque) is applied to the output shaft 38a (FIG. 9) on the basis of the slip generated in the toroidal-type continuously variable transmission 25 in a state where the brake pedal is depressed and stopped. Join. It is conceivable that the magnitude of this torque is set to a value commensurate with the creep force generated by a general automatic transmission equipped with a torque converter that has been widely used. This is because a driver who is used to operating a general automatic transmission does not feel uncomfortable. The direction of the torque is determined by the operation position of the operation lever provided at the driver's seat. When the operation lever is selected in the forward direction position (D range), torque is applied to the output shaft 38a in the forward direction, and when the reverse direction position (R range) is selected, the output lever 38 is moved in the reverse direction. Apply torque.

Next, a circuit for controlling the speed ratio of the toroidal type continuously variable transmission 25 in order to make the torque T _CVU1 actually passing through the toroidal type continuously variable transmission 25 coincide with the target value T _CVU2 as described above. The operation will be described with reference to FIG. Through a control valve 12, a pair of hydraulic chambers 24a and 24b constituting a hydraulic actuator 10 for displacing the trunnion 7 in the axial direction (vertical direction in FIG. 13) of the pivots 9 and 9 (see FIG. 7) Pressure oil can be supplied and discharged freely. The sleeve 14 constituting the control valve 12 is displaceable in the axial direction by the stepping motor 13 via the link arm 52 and the rod 53. The spool 15 constituting the control valve 12 is engaged with the trunnion 7 via a link arm 19, a recess cam 18 and a rod 17, and the axial displacement and swinging displacement of the trunnion 7 Displaceable in the direction. The above configuration is basically the same as a conventionally known transmission ratio control mechanism for a toroidal-type continuously variable transmission.

In particular, in the case of the prior invention, the sleeve 14 is driven not only by the stepping motor 13 but also by a hydraulic differential pressure cylinder 54. That is, the distal end portion of the rod 53 having the base end portion coupled to the sleeve 14 is pivotally supported at the intermediate portion of the link arm 52, and the stepping motor 13 is inserted into the long holes provided at both end portions of the link arm 52. Alternatively, a pin provided at the output portion of the differential pressure cylinder 54 is engaged. When a pin in a long hole provided at one end of the link arm 52 is pushed and pulled, the pin in the long hole at the other end serves as a fulcrum. With such a configuration, the sleeve 14 can be displaced in the axial direction not only by the stepping motor 13 but also by the differential pressure cylinder 54. In the case of the prior invention, the speed ratio e _CVU of the toroidal continuously variable transmission 25 is adjusted according to the torque T _CVU passing through the toroidal continuously variable transmission 25 by the displacement of the sleeve 14 by the differential pressure cylinder 54. I try to do it.

Therefore, in the case of the prior invention, different hydraulic pressures can be introduced into the pair of hydraulic chambers 55a, 55b provided in the differential pressure cylinder 54 through the correction control valve 56. The hydraulic pressure introduced into each of the hydraulic chambers 55a and 55b includes the differential pressure ΔP between the hydraulic pressures P _DOWN and P _UP acting in the pair of hydraulic chambers 24a and 24b constituting the actuator 10, and the correction control valve. It is determined based on the differential pressure ΔP ₀ between the output pressures of the pair of solenoid valves 57a, 57b for adjusting the opening degree of 56. That is, the opening and closing of both the solenoid valves 57a and 57b is performed so that the differential pressure ΔP ₀ between the output pressures of both the solenoid valves 57a and 57b is _equal to the target differential pressure corresponding to the target torque T _CVU2 of the toroidal continuously variable transmission 25 In this manner, the calculation is performed by a controller (controller) (not shown) and controlled based on an output signal output from the controller. Accordingly, the spool 58 constituting the correction control valve 56 has a force corresponding to the differential pressure ΔP of the hydraulic pressure acting in the hydraulic chambers 24a and 24b of the actuator 10 and a force against it. The differential pressure ΔP ₀ between the output pressures of the solenoid valves 57a and 57b, which is the target differential pressure corresponding to the target torque T _CVU2 , acts.

When the torque T _CVU1 that actually passes through the toroidal-type continuously variable transmission 25 and the target torque T _CVU2 coincide, that is, when the difference ΔT between the passing torque T _CVU1 and the target torque T _CVU2 is zero. The force according to the differential pressure ΔP of the hydraulic pressure acting in the hydraulic chambers 24a, 24b of the actuator 10 and the force according to the differential pressure ΔP ₀ of the output pressure of the electromagnetic valves 57a, 57b are balanced. Therefore, the spool 58 constituting the correction control valve 56 is in a neutral position, and the pressure acting on the hydraulic chambers 55a and 55b of the differential pressure cylinder 54 is also equal. In this state, the spool 59 of the differential pressure cylinder 54 is in the neutral position, and the speed ratio of the toroidal type continuously variable transmission 25 is not changed (not corrected).

On the other hand, when the difference between the torque T _CVU1 and the target torque T _CVU2 actually passing through the toroidal type continuously variable transmission 25 occurs, the hydraulic pressure chamber 24a of the actuator 10, the hydraulic pressure of the differential pressure acting on the 24b △ P And the force corresponding to the pressure difference ΔP ₀ between the output pressures of the electromagnetic valves 57a and 57b are lost. Then, the spool 58 constituting the correction control valve 56 is displaced in the axial direction according to the magnitude and direction of the difference ΔT between the passing torque T _CVU1 and the target torque T _CVU2, and the hydraulic pressure of the differential pressure cylinder 54 is changed. An appropriate hydraulic pressure is introduced into the chambers 55a and 55b according to the magnitude and direction of ΔT. Then, the spool 59 of the differential pressure cylinder 54 is displaced in the axial direction, and accordingly, the sleeve 14 constituting the control valve 12 is displaced in the axial direction. As a result, the trunnion 7 is displaced in the axial direction of the pivots 9 and 9, and the speed ratio of the toroidal continuously variable transmission 25 is changed (corrected). The direction in which the gear ratio changes and the amount of change are as described above with reference to FIGS. Further, the amount by which the speed ratio of the toroidal continuously variable transmission 25 is displaced, that is, the amount to be corrected (speed ratio correction amount) is sufficient with respect to the speed ratio width of the toroidal continuously variable transmission 25. It is a small one.

In order to realize the structure of the prior invention as described above, first, the torque T _CVU1 that actually passes through the toroidal-type continuously variable transmission 25 based on the hydraulic pressure acting in the hydraulic chambers 24a, 24b of the actuator 10. It is necessary to calculate accurately. Further, the hydraulic pressure corresponding to the magnitude and direction of the difference ΔT between the passing torque T _CVU1 and the target torque T _CVU2 is _strictly set in the hydraulic chambers 55a and 55b of the differential pressure cylinder 54 through the correction control valve 56. Need to be introduced in a regulated state. However, there are individual differences in the hydraulic pressure sensors for measuring the hydraulic pressure in the hydraulic chambers 24a and 24b. Unless the individual differences are grasped, the absolute value of the hydraulic pressure can be detected with high accuracy. Have difficulty. Further, even in the correction control valve 56 and the differential pressure cylinder 54, due to the pressure receiving area of the built-in spools 58, 59 and the variation in spring elasticity for returning the spools 58, 59 to the neutral position, etc. Subtle individual differences may occur in the operating state.

  In consideration of these matters, it is possible to accurately know individual characteristics of each component including the hydraulic sensor, the correction control valve 56, and the differential pressure cylinder 54. It becomes important to realize. Knowing the individual characteristics of each of these components is actually incorporating these components into a continuously variable transmission as shown in FIG. 8 or 9, and operating the continuously variable transmission and dynamo in combination. It is possible by things. However, it is troublesome to obtain the characteristics of each component in this way because many parts constituting the continuously variable transmission need to be assembled or disassembled each time a test is performed. And the operation | work which adjusts the characteristic of each said component becomes troublesome. Furthermore, if an excessive torque passes through the toroidal continuously variable transmission during operation with components having inappropriate characteristics incorporated, the toroidal continuously variable transmission may be damaged.

JP 2000-220719 A JP-A-10-103461

  In view of the circumstances as described above, the present invention provides a transmission ratio of a toroidal continuously variable transmission that constitutes a continuously variable transmission capable of realizing an infinite gear ratio by stopping an output shaft while rotating an input shaft. The present invention has been invented to realize a toroidal type continuously variable transmission test device that can easily determine the characteristics of the components of the gear ratio control unit for control.

The toroidal continuously variable transmission test apparatus of the present invention is a hydraulic actuator whose hydraulic pressure is controlled by a gear ratio control unit, and adjusts the gear ratio by displacing a support member that supports a power roller. This is for testing the performance of the transmission ratio control unit of the toroidal-type continuously variable transmission that performs the above.
Such a test apparatus for a toroidal type continuously variable transmission according to the present invention includes a first hydraulic cylinder unit, a second hydraulic cylinder unit, and a hydraulic control device.
Of these, the first hydraulic cylinder unit functions in the same manner as the actuator based on the supply and discharge of hydraulic pressure, and applies an axial force to the rod.
The second hydraulic cylinder unit applies an axial force to the rod based on supply and discharge of hydraulic pressure independent of the first hydraulic cylinder unit.
Further, the hydraulic control device controls supply / discharge of hydraulic pressure to / from the second hydraulic cylinder unit, and according to a force applied from the power roller to the support member during operation of the toroidal continuously variable transmission. The hydraulic pressure introduced into the second hydraulic cylinder unit is controlled so as to apply the applied force to the rod.

  When using the toroidal continuously variable transmission test apparatus according to the present invention configured as described above, a hydraulic pressure of a desired value is introduced into the second hydraulic cylinder unit by the hydraulic control apparatus, and the rod is In addition, a force having a predetermined magnitude directed in a predetermined direction with respect to the axial direction is applied. In this state, a thrust load similar to that corresponding to a force called 2Ft is applied to the first hydraulic cylinder unit via this rod when torque is transmitted between the input side disk and the output side disk. . A pressure difference corresponding to 2Ft is generated between the hydraulic pressures in the pair of hydraulic chambers constituting the first hydraulic cylinder unit.

  Therefore, the first hydraulic cylinder unit is regarded as an actuator constituting the toroidal continuously variable transmission, and a gear ratio constituting the toroidal continuously variable transmission such as a hydraulic sensor, a correction control valve, a differential pressure cylinder, etc. If the components of the control unit are assembled, the characteristics of these components can be measured. In this way, by providing the second hydraulic cylinder unit, it is possible to create the same situation as when a toroidal continuously variable transmission is operated without assembling the input side and output side disks and power rollers. . Therefore, not only the work for obtaining the characteristics of the above-mentioned components is facilitated, but also the members of the toroidal-type continuously variable transmission are not damaged along with the work for obtaining the characteristics.

Preferably, when the present invention is implemented, the direction and magnitude of the force applied to the rod by the second hydraulic cylinder unit can be arbitrarily adjusted by the hydraulic control device. By configuring in this manner, the output shaft is stopped while the input shaft is rotated, and the transmission ratio control unit of the toroidal continuously variable transmission incorporated in the continuously variable transmission that realizes an infinite gear ratio. The characteristics of each component are determined in accordance with the state of use of the continuously variable transmission.
Preferably, the transmission ratio control unit is provided with a pair of hydraulic devices arranged adjacent to each other with one partition plate interposed therebetween. Then, the oil passages formed in the partition plate are communicated with the ports of the hydraulic devices so that the hydraulic pressures at the port portions can be measured. With this configuration, the work for obtaining the characteristics of each hydraulic device can be performed efficiently.

FIG. 1 shows a first embodiment of the present invention. The toroidal continuously variable transmission test apparatus according to the present embodiment includes a first hydraulic cylinder unit 60, a second hydraulic cylinder unit 61, and a hydraulic control device 62.
Of these, the first hydraulic cylinder unit 60 is suitable for the gear ratio adjustment actuator 10 (see FIGS. 7 and 13) of the toroidal-type continuously variable transmission, and is the same as this actuator 10 based on the supply and discharge of hydraulic pressure. To apply an axial force to the rod 63.

  The state in which the hydraulic pressure is introduced into the pair of hydraulic chambers 24 a and 24 b provided in such a state that the piston 16 a is sandwiched in the first hydraulic cylinder unit 60 can be switched through the control valve 12. This control valve 12 is the same as the control valve 12 incorporated in the above-described conventional toroidal-type continuously variable transmission and the continuously variable transmission according to the previous invention, and a spool 15 is provided on the inner diameter side of the sleeve 14. It is fitted so as to be axially displaceable. Pressure oil sucked from the oil reservoir 64 and discharged by a pressure pump (not shown) is sent to the hydraulic chamber 24a (24b) of the hydraulic chambers 24a and 24b via the control valve 12. At this time, the hydraulic pressure in the other hydraulic chamber 24b (24a) is discharged through the control valve 12.

  An intermediate portion of the link arm 52 is pivotally supported at the distal end portion of the rod 53 having the base end portion coupled to the end portion of the sleeve 14 so as to be swingable and displaceable. Further, the end of the output rod 65 of the stepping motor 13 is disposed at one end (the upper end in FIG. 1) of the link arm 52, and the output rod 66 of the differential pressure cylinder 54 is disposed at the other end (the lower end in FIG. 1). The distal end portions are coupled so as to be swingable and slightly displaceable in the length direction of the link arm 52. With this configuration, the sleeve 14 can be driven to be displaced in the axial direction by one or both of the stepping motor 13 and the differential pressure cylinder 54. Further, the axial movement of the rod 63 can be transmitted to the spool 15 via the link arm 19a.

  A pair of hydraulic chambers 55a and 55b provided in the differential pressure cylinder 54 is passed through a correction control valve 56 (see FIG. 13; omitted in FIG. 1) that is switched by a pair of electromagnetic valves 57a and 57b. The desired hydraulic pressure can be introduced freely. The open / closed state of both solenoid valves 57a, 57b is detected by a hydraulic sensor (not shown) for detecting the hydraulic pressure in a pair of hydraulic chambers 24a, 24b provided in the first hydraulic cylinder unit 60. Control based on the value. Such an arrangement and the action of finely adjusting the axial displacement of the rod 63 by moving the sleeve 14 based on the hydraulic pressures of the hydraulic chambers 24a and 24b are shown in FIG. The transmission ratio control mechanism of the continuously variable transmission according to the previous invention is basically the same as the mechanism that controls the state of introduction of hydraulic pressure into the pair of hydraulic chambers 24a and 24b provided in the actuator 10. That is, by providing such a mechanism, the control valve 12, the stepping motor 13, the differential pressure cylinder 54, the correction control valve 56, and the both are included in the transmission ratio control mechanism of the continuously variable transmission. The characteristics of the electromagnetic valves 57a and 57b are determined in accordance with the operating state of the toroidal type continuously variable transmission.

  In this way, the rod 63 is provided with the rods 63 so that the characteristics of the constituent elements 12, 13, 54, 56, 57a, 57b are obtained in accordance with the state of being actually incorporated in the toroidal continuously variable transmission. The second hydraulic cylinder unit 61 can apply a thrust load similar to that applied when the toroidal continuously variable transmission is operated. The second hydraulic cylinder unit 61 is configured by fitting a piston 68 in a cylinder 67 so as to be oil-tight and displaceable in the axial direction (left-right direction in FIG. 1). And a pair of hydraulic chambers 69a and 69b. Further, a rod 70 having the piston 68 fixed to the intermediate portion thereof is coupled to the rod 63 of the first hydraulic cylinder unit 60. Accordingly, the rod 63 of the first hydraulic cylinder unit 60 is controlled in a predetermined direction by controlling the introduction state of the hydraulic pressure into the pair of hydraulic chambers 69a and 69b provided in the second hydraulic cylinder unit 61. A thrust load of a predetermined size facing is applied.

  Then, the hydraulic pressure control device 62 controls the introduction state of the hydraulic pressure into the pair of hydraulic chambers 69a and 69b. The hydraulic control device 62 includes a pair of electric hydraulic control valves 71a and 71b, a controller 72 incorporating a microcomputer for controlling the hydraulic control valves 71a and 71b, and the second hydraulic pressure control valve. A displacement sensor 73 that measures the amount of axial displacement of the rod 70 of the cylinder unit 61 and sends the measured value to the controller 72 is provided. Such a hydraulic control device 62 adjusts the opening direction and the opening degree of the hydraulic control valves 71a and 71b in response to a command from the controller 72, and connects the hydraulic source 74 and the pair of hydraulic chambers 69a and 69b. Switch the communication status. A thrust load having a desired magnitude is applied to the rod 63 via the rod 70 by appropriately regulating the difference in hydraulic pressure between the hydraulic chambers 69a and 69b.

  As described above, the thrust load applied to the rod 63 in this way is transmitted between the input side disk 2 and the output side disks 5 and 5a (see FIGS. 5 to 9) during the operation of the toroidal type continuously variable transmission. The value is the same as that according to (proportional to) the force called 2Ft applied according to the applied torque. That is, the controller 72 is commensurate with 2Ft applied to the actuator 10 (see FIG. 7) via the power roller 6 and the trunnion 7 as a support member in accordance with the operation of the toroidal continuously variable transmission (in the same direction and A force (in magnitude) is applied to the rod 63 via the rod 70. The direction of this force is determined by which one of the hydraulic chambers 69a and 69b is increased. Similarly, the magnitude is determined by the pressure difference between the hydraulic chambers 69a and 69b and the piston 68. Determined by the product of the pressure receiving area. For example, as shown in FIG. 2, when the pressure in the hydraulic chamber 69a of both the hydraulic chambers 69a and 69b is changed as shown by the solid line α, and the pressure in the hydraulic chamber 69b is changed as shown by the solid line β, respectively. The thrust load applied to the rod 63 via the rod 70 changes as indicated by a solid line γ. Further, when the force is applied, the displacement amount of the rod 63 is regulated while measuring the displacement amount of the rod 63 from the measured value of the displacement sensor 73. The amount of axial displacement of the rod 63 is commensurate with the amount of axial displacement of the pivots 9, 9 (see FIG. 7) of the trunnion 7 for the speed change operation of the toroidal type continuously variable transmission.

  According to the toroidal continuously variable transmission testing apparatus of the present embodiment configured as described above, the control valve 12, the stepping motor 13, the differential pressure cylinder 54, the correction control valve 56, and the pair of control valves. The characteristics of the electromagnetic valves 57a and 57b are obtained in accordance with the operating state of the toroidal type continuously variable transmission. That is, by assembling these components 12, 13, 54, 56, 57a, 57b in a state in which the axial displacement amount of the rod 63 of the first hydraulic cylinder unit 60 is controlled, The assembled state of 13, 54, 56, 57a, and 57b can be made to be in a state that matches the actual assembled state of the toroidal type continuously variable transmission. In this state, a desired value of hydraulic pressure is introduced into the pair of hydraulic chambers 68a and 69b provided in the second hydraulic cylinder unit 61 by the hydraulic control device 62, and a predetermined direction directed to the rod 63 in a predetermined direction. When the force of the magnitude is applied, the state where the second hydraulic cylinder 60 is placed is the same as the state where the actuator 10 is placed during the operation of the toroidal continuously variable transmission.

  Therefore, if the characteristics of the constituent elements 12, 13, 54, 56, 57a, and 57b are measured in this state, the characteristics can be accurately grasped. In this way, when determining the characteristics of each of the constituent elements 12, 13, 54, 56, 57a, 57b, the input side and output side disks 2, 5, 5a, and power rollers constituting the toroidal continuously variable transmission. There is no need to assemble 6, 6, trunnions 7, 7, etc. Therefore, not only the work for obtaining the characteristics is facilitated, but also the constituent members of the toroidal continuously variable transmission are not damaged in connection with the work for obtaining the characteristics.

  3 to 4 show a second embodiment of the present invention. In the case of the present embodiment, the differential pressure cylinder 54 and the correction control valve 56, each constituting a gear ratio control unit, each of which is a hydraulic device, are arranged adjacent to each other via a partition plate 75. Then, oil passages 76a to 76c formed on one side of the partition plate 75 (upper surface in FIGS. 3 to 4) are communicated with the ports 77a to 77c of the differential pressure cylinder 54, and the hydraulic pressures of the respective ports 77a to 77c are provided. Can be measured freely. Further, an oil passage (not shown) formed on the other surface of the partition plate 75 is connected to a port (not shown) of the correction control valve 56 so that the oil pressure at the port portion (not shown) can be measured. In the case of the present embodiment, such a configuration makes it possible to efficiently perform the work of obtaining the characteristics of the differential pressure cylinder 54 and the correction control valve 56. Other configurations and operations are the same as those in the first embodiment.

1 is a hydraulic circuit diagram showing a first embodiment of the present invention. The diagram which shows one example of the state which changes the hydraulic pressure in a 2nd hydraulic cylinder unit at the time of a test. The block diagram which shows Example 2 of this invention. Section A sectional drawing of FIG. Sectional drawing which shows an example of the toroidal type continuously variable transmission conventionally known. BB sectional drawing of FIG. CC sectional drawing. FIG. 6 is a schematic cross-sectional view showing an example of a conventionally known continuously variable transmission. FIG. 3 is a schematic cross-sectional view showing an example of a continuously variable transmission that controls a gear ratio by a control device according to a previous invention. The diagram which shows the relationship between the speed ratio of the toroidal type continuously variable transmission (CVU) incorporated in this continuously variable transmission, and the speed ratio as this whole continuously variable transmission (T / M). The diagram which shows the relationship between the rotational speed of an engine, and a torque in order to demonstrate the state which controls a gear ratio with the control apparatus which concerns on a prior invention. The diagram which shows the relationship between the torque and gear ratio which pass a toroidal type continuously variable transmission, and the gear ratio as the whole continuously variable transmission. The hydraulic circuit diagram which shows the mechanism for adjusting the gear ratio of the toroidal type continuously variable transmission which comprises the continuously variable transmission of prior invention.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 Input shaft 2 Input side disk 3 Ball spline 4 Output gear 5, 5a Output side disk 6 Power roller 7 Trunnion 8 Support shaft 9 Pivot 10 Actuator 11 Support plate 12 Control valve 13 Stepping motor 14 Sleeve 15 Spool 16, 16a Piston 17 Rod DESCRIPTION OF SYMBOLS 18 Precess cam 19, 19a Link arm 20 Synchronous cable 21 Cam surface 22 Drive shaft 23 Press device 24a, 24b Hydraulic chamber 25 Toroidal type continuously variable transmission 26, 26a Planetary gear type transmission 27, 27a Carrier 28a, 28b Planetary gear element 29 First transmission shaft 30a, 30b Sun gear 31 Second transmission shaft 32, 32a Hollow rotary shaft 33 Sun gear 34 Planetary gear element 35 Ring gear 36, 36a Second carrier 37a, 37b Planetary gear element 38, 38a Out Shaft 39, 39a Second ring gear 40, 40a Low speed clutch 41, 41a High speed clutch 42 First planetary gear 43 Second planetary gear 44a, 44b Planetary gear element 45a, 45b Planetary gear element 46 Transmission shaft 47 First One sun gear 48 Second sun gear 49 Ring gear 50 Third sun gear 51a, 51b Planetary gear element 52 Link arm 53 Rod 54 Differential pressure cylinder 55a, 55b Hydraulic chamber 56 Control valve 57a, 57b Solenoid valve 58 Spool 59 Spool 60 First hydraulic cylinder unit 61 Second hydraulic cylinder unit 62 Hydraulic control device 63 Rod 64 Oil reservoir 65 Output rod 66 Output rod 67 Cylinder 68 Piston 69a, 69b Hydraulic chamber 70 Rod 71a, 71b Hydraulic control valve 72 Controller 73 Displacement sensor 7 4 Hydraulic source 75 Partition plate 76a, 76b, 76c Oil passage 77a, 77b, 77c Port

Claims (3)

  The above transmission ratio of a toroidal continuously variable transmission that adjusts the transmission ratio by displacing the support member that supports the power roller with a hydraulic actuator whose hydraulic pressure introduction state is controlled by the transmission ratio control unit. In order to test the performance of the control unit, a first hydraulic cylinder unit that functions in the same manner as the actuator described above based on the supply and discharge of hydraulic pressure and applies an axial force to the rod, and the first hydraulic cylinder unit A second hydraulic cylinder unit that applies axial force to the rod based on independent hydraulic supply and discharge, and a hydraulic control device that controls supply and discharge of hydraulic pressure to the second hydraulic cylinder unit, The hydraulic control device applies to the rod a force corresponding to a force applied from the power roller to the support member during operation of the toroidal-type continuously variable transmission. Ku, controls the hydraulic pressure introduced into the second hydraulic cylinder unit, the toroidal type continuously variable transmission testing device.   The toroidal continuously variable transmission testing device according to claim 1, wherein the direction and magnitude of the force applied to the rod by the second hydraulic cylinder unit can be arbitrarily adjusted by the hydraulic control device. The transmission ratio control unit is provided with a pair of hydraulic devices arranged adjacent to each other with one partition plate interposed between them, and the oil passage formed in the partition plate is connected to the ports of these hydraulic devices, The testing device for a toroidal type continuously variable transmission according to any one of claims 1 to 2, wherein the hydraulic pressure of each of the port portions can be measured.

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