JP2007032793A - Toroidal continuously variable transmission - Google Patents

Abstract

<P>PROBLEM TO BE SOLVED: To realize construction which can swiftly and accurately pick up difference of hydraulic pressure (differential pressure) existing between a respective pair of hydraulic pressure chambers 36a, 36b which is provided with an actuator 13 and moreover can construct to miniaturization at low cost. <P>SOLUTION: A toroidal continuously transmission picks up hydraulic pressure rising at first and second hydraulic introduction passages 54, 56 of a differential pressure pick-up valve 37 through one of pick-up oil passage 39. Therefore, it connects the first and second hydraulic introduction passages 54, 56 with each other via a ball type directional control valve 40. <P>COPYRIGHT: (C)2007,JPO&INPIT

Description

  The present invention relates to an improvement in a toroidal continuously variable transmission that is used as a transmission unit of an automatic transmission for automobiles or as a transmission for adjusting the operating speed of various industrial machines such as pumps.

  The use of a toroidal type continuously variable transmission as an automobile transmission is described in many publications such as Patent Documents 1 and 2 and Non-Patent Documents 1 and 2, and has been well-known in some implementations. is there. Also, continuously variable transmissions that combine a toroidal type continuously variable transmission and a planetary gear type transmission that is a differential unit in order to increase the fluctuation range of the gear ratio are described in, for example, Patent Documents 3 to 7, etc. Therefore, it has been widely known. FIGS. 4 to 5 are continuously variable transmissions having a mode that can realize a so-called geared neutral state described in Patent Documents 6 to 7, in which the output shaft can be stopped while the input shaft is rotated in one direction. The device is shown. 4 shows a block diagram of the continuously variable transmission, and FIG. 5 shows a hydraulic circuit for controlling the continuously variable transmission.

  The output of the engine 1 is input to the input shaft 3 via the damper 2. The power transmitted to the input shaft 3 is transmitted to the planetary gear type transmission 5 which is a differential unit, either directly or via the toroidal continuously variable transmission 4. The differential components of the constituent members of the planetary gear type transmission 5 are taken out to the output shaft 9 via the clutch device 6, that is, the low speed and high speed clutches 7 and 8 shown in FIG. The toroidal-type continuously variable transmission 4 includes a plurality of input and output disks 10 and 11, each of which is a first and second disk, a plurality of power rollers 12, and a plurality of support members. Each trunnion (not shown), an actuator 13 (FIG. 5), a pressing device 14, and a transmission ratio control unit 15 are provided. Of these, the input-side and output-side disks 10 and 11 are arranged concentrically and relatively freely rotatable. Each of the power rollers 12 is sandwiched between the inner surfaces of the input and output disks 10 and 11 facing each other, and the power roller 12 is driven between the input and output disks 10 and 11. To communicate. Each trunnion supports each power roller 12 rotatably.

  The actuator 13 is of a hydraulic type, and the trunnions supporting the power rollers 12 are displaced in the axial directions of the pivots provided at both ends so that the input side disk 10 and the output side The gear ratio with the disk 11 is changed. The pressing device 14 is of a hydraulic type and presses the input side disk 10 and the output side disk 11 in a direction approaching each other. The gear ratio control unit 15 controls the displacement direction and the displacement amount of the actuator 13 so that the gear ratio between the input side disk 10 and the output side disk 11 becomes a desired value.

  In the case of the illustrated example, the transmission ratio control unit 15 includes a controller 16, a stepping motor 17 that is switched based on a control signal from the controller 16, a line pressure control electromagnetic on-off valve 18, and an electromagnetic valve. 19, a shift electromagnetic valve 20, and a control valve device 21 whose operation state can be switched by these members 17 to 20. The control valve device 21 includes a transmission ratio control valve 22, a differential pressure cylinder 23, correction control valves 24a and 24b, and high-speed clutch and low-speed clutch switching valves 25 and 26 (FIG. 5). It is a combination. Of these, the gear ratio control valve 22 controls the supply and discharge of hydraulic pressure to the actuator 13. Further, the differential pressure cylinder 23 is configured to control the transmission ratio control valve so as to correct the transmission ratio of the toroidal type continuously variable transmission 4 according to the torque (passing torque) passing through the toroidal type continuously variable transmission 4. This is for adjusting the switching state of 22. The correction control valves 24 a and 24 b control the supply and discharge of pressure oil to and from the differential pressure cylinder 23. Further, the switching valves 25 and 26 for the high speed clutch and the low speed clutch switch the introduction state of the pressure oil to the low speed and high speed clutches 7 and 8, respectively.

  Further, the pressure oil discharged from the oil pump 27 (27a, 27b in FIG. 5) driven by the power extracted from the damper 2 is sent to the control valve device 21, the pressing device 14, and the like. That is, the pressure oil sucked from the oil reservoir 28 (FIG. 5) and discharged by the oil pumps 27a and 27b can be adjusted to a predetermined pressure by the pressing force adjusting valve 29 and the low pressure side adjusting valve 30 (FIG. 5). It is said. Of these adjusting valves 29 and 30, the pressing force adjusting valve 29 for adjusting the hydraulic pressure sent to the pressing device 14 and the manual hydraulic pressure switching valve 31 is described in detail in, for example, Patent Document 8 and the like. In addition, it has a function as a relief valve, and includes first to third pilot portions 32 to 34. Among these, the first and second pilot parts 32 and 33 open the pressing force adjusting valve 29 according to the magnitude of the force transmitted between the input side disk 10 and the output side disk 11. It is for adjusting the pressure. For this purpose, there exists between a pair of hydraulic chambers 36a and 36b provided with a piston 35 sandwiched between an actuator 13 for displacing a support member (trunnion) for supporting the power roller 12 in the axial direction of the pivot axis. The hydraulic pressure difference (differential pressure) is introduced into the first and second pilot parts 32 and 33 via the differential pressure take-out valve 37.

  On the other hand, the third pilot portion 34 is a transmission ratio of the toroidal type continuously variable transmission 4, the temperature of the lubricating oil (traction oil) existing in the toroidal type continuously variable transmission 4, and a drive source. This is for adjusting the valve opening pressure of the pressing force adjusting valve 29 in accordance with operating conditions other than the transmitted force such as the rotational speed of an engine 1. Therefore, based on the opening / closing (duty ratio control) of the line pressure control electromagnetic on / off valve 18 controlled by a command from the controller 16, pressure oil of a predetermined pressure can be introduced into the third pilot section 34 freely. Yes. Then, by appropriately adjusting the hydraulic pressure introduced into the first to third pilot portions 32 to 34, the pressing force generated by the pressing device 14 is set in accordance with the operating state of the toroidal continuously variable transmission 4. It is configured to properly regulate.

  Further, the pressure oil adjusted by the pressing force adjusting valve 29 passes through the manual hydraulic pressure switching valve 31, the high speed clutch switching valve 25 or the low speed clutch switching valve 26, and the low speed clutch 7 or the high speed clutch. The clutch 8 can be fed into the hydraulic chamber. The low speed clutch 7 out of the low speed and high speed clutches 7 and 8 is connected when realizing a low speed mode in which the reduction ratio is increased (including an infinite gear ratio (including a geared neutral state)). At the same time, the connection is broken when the high speed mode for reducing the reduction ratio is realized. In contrast, the high speed clutch 8 is disconnected when realizing the low speed mode and is connected when realizing the high speed mode. Further, the supply / discharge state of the pressure oil to the low speed and high speed clutches 7 and 8 is switched according to the switching of the shift solenoid valve 20.

  By the way, in the case of the conventional structure as shown in FIGS. 4 to 5 described above, the pressure difference between the pair of hydraulic chambers 36 a and 36 b constituting the actuator 13 is introduced into the pressing force adjusting valve 29. For this purpose, the pressure adjusting valve 29 and the differential pressure extracting valve 37 are connected by two oil passages 38a and 38b. In the case of the structure using two oil passages 38a and 38b as described above, there is a possibility that the valve body provided with each of the oil passages 38a and 38b may be larger than the structure in which this oil passage is made one. There is. In addition, the design and manufacturing work of the arrangement of these two oil passages 38a and 38b becomes troublesome and the cost may increase. In addition, in order to introduce hydraulic pressure to the first and second pilot portions 32 and 33 of the pressing force adjusting valve 29 through these two oil passages 38a and 38b, respectively, the structure has one oil passage. In comparison with this, the number of pilot parts and spools (valves) is increased, the structure of the pressing force adjusting valve 29 is complicated, and the pressing force adjusting valve 29 may be enlarged.

On the other hand, in Patent Document 9, hydraulic sensors are provided in a pair of hydraulic chambers constituting an actuator, and based on the hydraulic pressures (P _high , P _low ) of the hydraulic chambers detected from these hydraulic sensors, A technique is described in which a controller calculates a hydraulic pressure difference (differential pressure: ΔP = P _high −P _low ) existing between the hydraulic chambers. In addition, Patent Document 9 describes a technique for adjusting the driving force (creep force) output from the output shaft of the continuously variable transmission based on the differential pressure calculated by the controller in this way. Has been. If such a technique described in Patent Document 9 is adopted, the differential pressure between a pair of hydraulic chambers constituting the actuator can be obtained without using the differential pressure take-out valve 37 as described above. Also, if the hydraulic pressure introduced into the pressing device can be adjusted based on the differential pressure calculated by such a controller, the number of oil passages can be reduced compared to the structure using the differential pressure take-out valve 37 described above. It is considered possible.

  However, in the case of the conventional technique described in Patent Document 9, there is a possibility that an error is included in the differential pressure calculated by the controller (deviates from the actual value, becomes inaccurate) depending on the operating condition. There is. Hereinafter, this point will be described. That is, the value of the hydraulic pressure in each hydraulic chamber constituting the actuator fluctuates finely (vibrates) when, for example, the torque passing through the toroidal continuously variable transmission fluctuates suddenly. In such a case, for example, the hydraulic pressure of the hydraulic chamber having the higher hydraulic pressure among the pair of hydraulic chambers is detected as the largest value in the state where the hydraulic pressure is finely raised and lowered, and the hydraulic chamber having the lower hydraulic pressure is also detected. If the hydraulic pressure is detected at the smallest value in the same finely rising and falling state, the differential pressure calculated by the controller may be considerably larger than the actual value. On the contrary, the hydraulic pressure in the hydraulic chamber on the higher hydraulic pressure side is detected with the smallest value in the state where the hydraulic chamber is finely raised and lowered, and the hydraulic pressure in the lower hydraulic chamber is also finely raised and lowered. If the maximum value is detected, the differential pressure calculated by the controller may be considerably smaller than the actual value.

  In such a case, for example, if the differential pressure calculated by the controller is used as it is for adjusting the driving force (creep force) output from the output shaft of the continuously variable transmission, the driving force required from the output shaft is used. May not be output. Further, when the pressing force of the pressing device is adjusted based on the differential pressure calculated by the controller, when the pressing force becomes excessively small, a slip (gross slip) occurs at the rolling contact portion, or the pressing force is reversed. When the pressure becomes excessively large, transmission efficiency and durability may be reduced. In order to prevent such inconvenience, it is conceivable to remove vibration from the detection signals of the respective hydraulic sensors by filtering the detection signals of the respective hydraulic sensors (for example, by a low-pass filter).

  However, in such a case, there is a possibility that this detection signal is output later than the actual detection time by the amount of filtering of the detection signal. Such a delay in the detection signal hardly causes a problem if the hydraulic pressure is stable {if it is not finely oscillated (vibrated)}. However, for example, when the torque passing through the toroidal type continuously variable transmission fluctuates due to a sudden depression of an accelerator pedal or the like, and when the hydraulic pressure of each hydraulic chamber of the actuator suddenly fluctuates based on the fluctuation of the torque, The differential pressure calculated by the controller may deviate from the actual value (cannot catch up with the actual value). In spite of the deviation from the actual value in this way, if the pressing force of the pressing device is adjusted, for example, based on the differential pressure calculated by the controller, the pressing force required from the pressing device May not be generated quickly. For example, when the accelerator pedal is depressed suddenly as described above, it takes time until the required pressing force is generated, during which time slippage occurs at the rolling contact portion, and transmission efficiency and durability may be reduced. There is sex. Similarly, when adjusting the driving force output from the output shaft of the continuously variable transmission, it takes time to output the required driving force, which is also not preferable.

Japanese Patent No. 2734583 JP-A-5-39850 Japanese Patent Laid-Open No. 10-196759 JP 2003-307266 A JP 2000-220719 A JP 2004-225888 A JP 2004-211836 A JP 2004-76940 A JP 2002-213591 A Motoo Aoyama, "Bessed Best Car Red Badge Series 245 / A book that understands the latest mechanics of cars", Sanyusha Co., Ltd./Kodansha Co., Ltd., December 20, 2001, p. 92-93 Hirohisa Tanaka, “Toroidal CVT”, Corona Inc., July 13, 2000

  In view of the circumstances as described above, the present invention has a structure that can be made small and inexpensive, and that can quickly and accurately take out the differential pressure between a pair of hydraulic chambers provided with a piston sandwiched between hydraulic actuators. It was invented to realize.

The toroidal type continuously variable transmission of the present invention is similar to the conventionally known toroidal type continuously variable transmissions, and includes first and second disks, a plurality of support members, a plurality of power rollers, and an actuator. And a differential pressure take-off valve.
Of these, the first and second disks are arranged concentrically and relatively rotatably.
Each of the supporting members is supported so as to be swingable and displaceable about a pivot that is twisted with respect to the central axis of the two disks.
The power rollers are sandwiched between the inner surfaces of the first and second disks facing each other in a state of being rotatably supported by the support members.
The actuator is a hydraulic actuator, and the piston is displaced in the axial direction based on the supply and discharge of the pressure oil to and from the first and second hydraulic chambers provided across the piston, and the support members are It is displaced in the axial direction of the pivot.
The differential pressure take-off valve takes out a hydraulic pressure proportional to the difference between the hydraulic pressures existing between the first and second hydraulic chambers constituting one of the actuators.
In particular, in the toroidal type continuously variable transmission according to the present invention, the absolute difference in hydraulic pressure existing between the first and second hydraulic chambers through the one take-out oil passage from the differential pressure take-out valve is absolute. Take out the hydraulic pressure proportional to the value.

  According to the toroidal continuously variable transmission of the present invention configured as described above, it is proportional to the absolute value of the difference in hydraulic pressure (differential pressure) existing between the first and second hydraulic chambers constituting the actuator. Since oil pressure is taken out from the differential pressure take-out valve through one take-out oil passage, the oil passage arrangement design and the oil pressure responsiveness (the degree of responsiveness and follow-up is good) for taking out the differential pressure are ensured. The manufacturing operation can be facilitated, and further, the valve body provided with this oil passage can be reduced in size and cost. Further, if the differential pressure is introduced into, for example, a pressure adjusting valve for adjusting the pressure to be introduced into the pressing device through such a single take-out oil passage, the pilot portion and the spool of the pressure adjusting valve are connected. Less. For this reason, the pressing force adjusting valve can be configured in a simple and compact manner, and also from this aspect, the valve body incorporating the pressing force adjusting valve can be reduced in size and cost. Even when the differential pressure is detected by a hydraulic sensor, the differential pressure can be detected as it is by providing a single hydraulic sensor in the take-out oil passage. For this reason, as described above, there is an error in the calculated value as in the case where the hydraulic pressure in each of the first and second hydraulic chambers is detected by a pair of hydraulic pressure sensors and the differential pressure is calculated from the hydraulic pressure in each hydraulic chamber. It can also be prevented from being included. Furthermore, it is not necessary to filter the detection signal of the hydraulic sensor, and the differential pressure can be obtained quickly and accurately.

In carrying out the present invention, preferably, as described in claim 2, the differential pressure take-out valve includes a spool (one or more), a source hydraulic pressure introduction port, and first and second differential pressure feeds. Port.
Of these, the spool is displaced in accordance with a difference in hydraulic pressure (differential pressure) existing between the first and second hydraulic chambers constituting the actuator. The source oil pressure introduction port is for introducing the oil pressure generated by the oil pressure source. The first differential pressure feed port communicates with the source hydraulic pressure introduction port when the hydraulic pressure in the first hydraulic chamber is higher than the hydraulic pressure in the second hydraulic chamber. The second differential pressure delivery port communicates with the source hydraulic pressure introduction port when the hydraulic pressure in the second hydraulic chamber is higher than the hydraulic pressure in the first hydraulic chamber. Of the first and second differential pressure delivery ports, a port communicating with the source hydraulic pressure introduction port is connected to the take-out oil passage so as to be freely communicated according to the difference in hydraulic pressure at that time. For this purpose, for example, as described in claim 3, the first and second differential pressure delivery ports and the take-out oil passage are connected via a ball-type direction switching valve (for example, a shuttle valve).
With this configuration, any one of the first and second differential pressure delivery ports that communicates with the source hydraulic pressure introduction port according to the hydraulic pressure difference (differential pressure) and the take-out oil The road can be properly communicated with the direction switching valve. Therefore, the hydraulic pressure proportional to the absolute value of the hydraulic pressure difference existing between the first and second hydraulic chambers can be taken out with good responsiveness through the differential pressure take-off valve, and the differential pressure can be It can be obtained quickly and accurately.

More preferably, as described in claim 4, according to the oil pressure of the take-out oil passage detected by the oil pressure sensor, the transmission ratio between the first and second disks, and the first and second At least one of the pressing force of the pressing device that presses the disks toward each other is adjustable.
If comprised in this way, a differential pressure | voltage can be detected rapidly and accurately with one hydraulic pressure sensor provided in the extraction oil path. For this reason, if the gear ratio and the pressing force are adjusted according to the differential pressure, the necessary gear ratio and the pressing force can always be obtained reliably.

  FIGS. 1-2 has shown Example 1 of this invention corresponding to Claims 1-3. The feature of this embodiment is that an absolute difference in pressure (differential pressure) between a pair of (first and second) hydraulic chambers 36a and 36b provided on the actuator 13 with a piston 35 interposed therebetween is absolute. The oil pressure proportional to the value is taken out through one take-out oil passage 39. Since the structure and operation of the other parts are the same as those of the conventional structure shown in FIGS. 4 to 5 described above, overlapping illustrations and descriptions are omitted or simplified, and the following description will focus on the characteristic parts of this embodiment. .

  As shown in detail in FIG. 2, the differential pressure take-out valve 37 of the present embodiment incorporates a spool 42 in a cylinder hole 41 so as to be freely displaceable in the axial direction (left and right direction in FIGS. 1 and 2). Further, a source oil pressure introduction port 43 (FIG. 2) is provided at the center of the cylinder hole 41. The source oil pressure introduction port 43 is connected to a discharge port of an oil pump 27a, which is a hydraulic power source, via a pressure introduction path 44 provided with a pressing force adjusting valve 29a on the way. Accordingly, the oil pressure discharged from the oil pump 27a and adjusted by the pressing force adjusting valve 29a is introduced into the central portion of the cylinder hole 41 through the source oil pressure introduction port 43.

  Further, a first differential pressure delivery port 45 and a second differential pressure delivery port 46 (FIG. 2) are provided at portions sandwiching the source hydraulic pressure introduction port 43 from both sides in the axial direction and slightly closer to both ends from the center of the cylinder hole 41. And are provided. Of these, the first differential pressure delivery port 45 is connected to the source oil pressure introduction port 43 in a state where the spool 42 is displaced from the neutral position shown in FIGS. 1-2 to one side in the axial direction (left side in FIGS. 1-2). Communicate. On the other hand, the second differential pressure delivery port 46 communicates with the source hydraulic pressure introduction port 43 in a state where the spool 42 is displaced from the neutral position to the other axial side (the right side in FIGS. 1 and 2).

  Further, drain ports 47a and 47b (FIG. 2) are provided in the intermediate portion of the cylinder hole 41 at the portion sandwiching the first and second differential pressure delivery ports 45 and 46 from both sides in the axial direction. Both the drain ports 47a and 47b communicate with an oil reservoir 28 such as an oil pan. Accordingly, there is no hydraulic pressure in the drain ports 47a and 47b (gauge pressure = 0). Further, a first hydraulic pressure introduction port 48 and a second hydraulic pressure introduction port 49 are provided in a portion sandwiching both the drain ports 47a and 47b from both sides in the axial direction, which are closer to both ends in the middle portion of the cylinder hole 41. . The hydraulic pressure in one hydraulic chamber 36a of the actuator 13 for displacing the trunnion as a support member is introduced into the first hydraulic pressure introduction port 48 through the second pressure introduction passage 50a. In contrast, the hydraulic pressure in the other hydraulic chamber 36b of the actuator 13 is introduced into the second hydraulic pressure introduction port 49 through the second pressure introduction path 50b.

  Further, a first reaction force port 51 and a second reaction force port 52 (FIG. 2) are provided in the vicinity of both end portions in the axial direction of the cylinder hole 41. Of these, the first reaction force port 51 communicates with a first reaction force chamber 53 provided at one end (the left end in FIGS. 1 and 2) of the cylinder hole 41. When the hydraulic pressure is introduced into the first reaction force chamber 53, the spool 42 receives a rightward force in FIGS. 1 and 2 corresponding to the product of the hydraulic pressure and the cross-sectional area of the first reaction force chamber 53. Further, the first reaction force port 51 and the first differential pressure delivery port 45 are communicated with each other by a first hydraulic pressure introduction passage 54, and the hydraulic pressure at the first differential pressure delivery port 45 is supplied to the first reaction force chamber. 53.

  The second reaction force port 52 communicates with a second reaction force chamber 55 provided at the other end of the cylinder hole 41 (the right end in FIGS. 1 and 2). When the hydraulic pressure is introduced into the second reaction force chamber 55, the spool 42 receives a leftward force in FIGS. 1 and 2 corresponding to the product of the hydraulic pressure and the cross-sectional area of the second reaction force chamber 55. Further, the second reaction force port 52 and the second differential pressure delivery port 46 are communicated with each other by a second hydraulic pressure introduction passage 56, and the hydraulic pressure at the second differential pressure delivery port 46 is supplied to the second reaction force chamber. 55 is introduced. The spool 42 is elastically pressed toward the neutral position by a pair of return springs 57, 57 installed in the first and second reaction force chambers 53, 55. Accordingly, the axial position of the spool 42 is the neutral position shown in FIGS. 1 and 2 as long as the hydraulic pressures of the first and second hydraulic pressure introduction ports 48 and 49 are equal to each other.

  The differential pressure take-out valve 37 configured as described above does not transmit torque between the input side and output side disks 10 and 11 (see FIG. 4) constituting the toroidal-type continuously variable transmission 4, and the both When the hydraulic pressures in the hydraulic chambers 36a and 36b are equal to each other, the spool 42 is in a neutral position as shown in FIGS. 1 and 2 due to the balance between the return springs 57 and 57. In this state, the first and second differential pressure feed ports 45 and 46 and the source hydraulic pressure introduction port 43 are shut off. For this reason, the oil pressure at the source oil pressure introduction port 43 is not sent to the differential pressure feed ports 45 and 46.

  On the other hand, when torque is transmitted between the disks 10 and 11, and there is a difference in the hydraulic pressure in the hydraulic chambers 36a and 36b, the first and second hydraulic pressure introduction ports are used. Different hydraulic pressures are introduced into the 48 and 49 portions. Then, the spool 42 moves in the axial direction against the elasticity of the return springs 57 and 57 based on the difference between the hydraulic pressures introduced into the hydraulic pressure introduction ports 48 and 49. For example, when the hydraulic pressure in one hydraulic chamber (first hydraulic chamber) 36a among the hydraulic chambers 36a and 36b is higher than the hydraulic pressure in the other hydraulic chamber (second hydraulic chamber) 36b, The spool 42 is displaced to the left in the figure, and the source hydraulic pressure introduction port 43 and the first differential pressure feed port 45 communicate with each other. The hydraulic pressure rises at the first differential pressure delivery port 45 portion. At this time, the spool 42 moves to the axial position determined by the balance of the product of the oil pressure and the pressure receiving area of each part, so that the oil pressure in the first differential pressure delivery port 45 portion is the both hydraulic chambers 36a, 36b. The value is proportional to the hydraulic pressure difference.

  On the other hand, when the hydraulic pressure in the other hydraulic chamber (second hydraulic chamber) 36b among the hydraulic chambers 36a and 36b is higher than the hydraulic pressure in the same hydraulic chamber (first hydraulic chamber) 36a, The spool 42 is displaced to the right in the figure, and the source hydraulic pressure introduction port 43 and the second differential pressure feed port 46 communicate with each other. The hydraulic pressure rises at the second differential pressure feed port 46 portion. Also in this case, the spool 42 moves to the axial position determined by the balance between the product of the oil pressure and the pressure receiving area of each part, so that the oil pressure in the second differential pressure delivery port 46 portion is the both hydraulic chambers 36a, The value is proportional to the hydraulic pressure difference in 36b.

  Further, in the case of this embodiment, as described above, the hydraulic pressure rising at the first and second differential pressure delivery ports 45 and 46 of the differential pressure relief valve 37 {a pair of hydraulic chambers 36a and 36b constituting the actuator 13 The hydraulic pressure proportional to the absolute value of the hydraulic pressure difference (differential pressure) existing between the two oil pressures is freely taken out through one take-out oil passage 39. For this purpose, of the first and second differential pressure delivery ports 45, 46, the differential pressure delivery port 45 (46) communicating with the source hydraulic pressure introduction port 43 in accordance with the differential pressure at that time is removed. The first and second differential pressure feed ports 45 and 46 and the take-out oil passage 39 are connected via a ball-type direction switching valve 40 so as to communicate with the oil passage 39. That is, a pair of the first hydraulic pressure introduction path 54 connected to the first differential pressure delivery port 45 and the second hydraulic pressure introduction path 56 connected to the second differential pressure delivery port 46 are provided in the direction switching valve 40. Connected to the input port. And, according to the switching of the direction switching valve 40, the hydraulic pressure introduction path communicating with the source hydraulic pressure introduction port 43 (the pressure introduction path 44 connected to) of the first and second hydraulic pressure introduction paths 54, 56. 54 (56) is communicated with the take-out oil passage 39 connected to an output port provided only for the direction switching valve 40. Such a direction switching valve 40 is formed by fitting a check ball 58 in a cylinder hole 59.

  For example, when the oil pressure of the source oil pressure introduction port 43 is introduced into the first oil pressure introduction path 54 (the spool 42 is displaced to the left in FIG. 2), the check ball 58 of the direction switching valve 40 is based on the oil pressure. However, it is displaced to the right in FIG. As a result, the second hydraulic pressure introduction path 56 is blocked by the check ball 58, the first hydraulic pressure introduction path 54 communicates with the take-out oil path 39, and the differential pressure (absolute Value) is introduced. On the contrary, when the oil pressure of the source oil pressure introduction port 43 is introduced into the second oil pressure introduction path 56 (the spool 42 is displaced to the right in FIG. 2), the direction switching valve 40 is based on this oil pressure. The check ball 58 is displaced to the left in FIG. As a result, the first hydraulic pressure introduction path 54 is blocked by the check ball 58, and the second hydraulic pressure introduction path 56 communicates with the take-out oil path 39. ) Is introduced in proportion to the hydraulic pressure.

  On the other hand, when the hydraulic pressure is not introduced into the first and second hydraulic pressure introduction paths 54 and 56 (the differential pressure is 0) (the spool 42 is in the neutral position), the check ball 58 is in the neutral position and the first The second pressure introducing passages 54 and 56 and the take-out oil passage 39 are blocked. Therefore, in the case of the present embodiment, one of the first and second hydraulic pressure introduction paths 54, 56 is based on the differential pressure existing between the hydraulic chambers 36a, 36b. If an oil pressure proportional to the differential pressure is introduced to (56), this oil pressure is introduced into the take-out oil passage 39 regardless of which oil pressure introduction path 54 (56) is introduced. .

  Further, in the case of the present embodiment, the pressure difference adjusting valve 29a for adjusting the hydraulic pressure sent to the pressing device 14 through the (one) take-out oil passage 39 is set to the above-described differential pressure (absolute value). Proportional hydraulic pressure is introduced. The pressing force adjusting valve 29 having the conventional structure shown in FIG. 5 described above has three (first to third) pilot portions 32 to 34, whereas the pressing force adjusting valve of this embodiment is used. In the case of 29a, two (first and second) pilot sections 32a and 33a are provided. Then, by introducing a hydraulic pressure proportional to the differential pressure (absolute value) through the take-out oil passage 39 into the first pilot portion 32a, the input side disk 10 and the output side disk 11 The valve opening pressure of the pressing force adjusting valve 29a is adjusted according to the magnitude of the force transmitted between them. Specifically, the higher the hydraulic pressure introduced into the take-out oil passage 39, the higher the valve opening pressure of the pressing force adjusting valve 29a, and the higher the hydraulic pressure introduced into the pressing device 14. On the other hand, the second pilot portion 33a includes a gear ratio of the toroidal-type continuously variable transmission 4, the temperature of lubricating oil (traction oil) existing in the toroidal-type continuously variable transmission 4, and an engine serving as a drive source. 1 is controlled by a command from the controller 16 (see FIG. 4) in order to adjust (depressurize) the valve opening pressure of the pressing force adjusting valve 29a in accordance with operating conditions other than the transmitted force such as the rotational speed of 1. The pressure oil based on the opening / closing (duty ratio control) of the line pressure control electromagnetic switching valve 18 is introduced. Then, by appropriately adjusting the hydraulic pressure introduced into the first and second pilot portions 32a and 33a in this way, the pressing force generated by the pressing device 14 is changed to the operating state of the toroidal continuously variable transmission 4. Appropriate regulation.

  As described above, in this embodiment, the hydraulic pressure proportional to the absolute value of the hydraulic pressure difference (differential pressure) existing between the pair of hydraulic chambers 36a and 36b constituting the actuator 13 is extracted. It is taken out from the valve 37 through one take-out oil passage 39. For this reason, while ensuring the hydraulic responsiveness regarding the extraction of the differential pressure, the arrangement design and manufacturing work of the extraction oil passage 39 are facilitated, and further, the valve body provided with the extraction oil passage 39 is reduced in size and cost. Can be planned. Further, since the differential pressure is introduced into the pressing force adjusting valve 29a through such one take-out oil passage 39, the pilot portions 32a and 33a of the pressing force adjusting valve 29a can be reduced. For this reason, the pressing force adjusting valve 29a can be configured in a simple and compact manner, and also from this aspect, the valve body incorporating the pressing force adjusting valve 29a can be reduced in size and cost. The differential pressure take-out valve 37 is not limited to the structure shown in this embodiment. The hydraulic pressure is one or more spools, and the hydraulic pressure proportional to the absolute value of the hydraulic pressure difference (differential pressure) existing between the pair of hydraulic chambers 36a, 36b constituting the actuator 13 is 1 Any one can be adopted as long as it can be taken out through a take-out oil passage (take-out port).

  In the case of the present embodiment, the structure of the transmission ratio control unit 15 for adjusting the transmission ratio of the toroidal type continuously variable transmission 4 is also improved in order to improve the characteristics at the start of the vehicle (characteristic in the geared neutral state). Devised. That is, in the case of the conventional structure shown in FIGS. 4 to 5 described above, the gear ratio of the toroidal continuously variable transmission 4 is adjusted (corrected) by the differential pressure cylinder 23 in addition to the stepping motor 17 when the vehicle starts. Thus, the torque passing through the toroidal-type continuously variable transmission 4 and thus the driving force (creep force) output from the output shaft 9 is controlled (torque control). That is, the sleeve 60 of the transmission ratio control valve 22 is finely adjusted by the differential pressure cylinder 23 in addition to being displaced in the axial direction by the stepping motor 13. On the other hand, in the case of the present embodiment, the hydraulic second actuator 61 is driven in accordance with the switching of the manual hydraulic pressure switching valve 31 based on the operation of the shift lever at the driver's seat, and the speed ratio control valve 22 is controlled. By displacing the sleeve 60, a predetermined driving force can be output from the output shaft 9 (see FIG. 4) when the vehicle starts.

  That is, when the shift lever is operated to a parking position (P range), a neutral position (N range), or a forward position (D, L range), the second hydraulic pressure switching valve 31 is changed based on the switching of the manual hydraulic pressure switching valve 31. The hydraulic chamber 62 constituting the actuator 61 communicates with the oil reservoir 28, and the pressure oil in the hydraulic chamber 62 is discharged. As a result, the spool 63 constituting the second actuator 61 is displaced in the other axial direction (leftward in FIG. 1) based on the elasticity of the spring 64. Based on the displacement of the spool 63, the sleeve 60 of the gear ratio control valve 22 is displaced in the other axial direction, and the gear ratio of the toroidal continuously variable transmission 4 (see FIG. 4) is corrected by a predetermined amount. On the other hand, when the shift lever is operated to the reverse position (R range), pressure oil is supplied to the hydraulic chamber 61 based on switching of the manual hydraulic pressure switching valve 31. As a result, the spool 63 of the second actuator 61 is moved in the axial direction (to the right in FIG. 1) based on the supply of pressure oil into the hydraulic chamber 62 (against the elasticity of the spring 64). Displace. Based on the displacement of the spool 63, the sleeve 60 of the transmission ratio control valve 22 is displaced in one axial direction, and the transmission ratio of the toroidal continuously variable transmission 4 is corrected by a predetermined amount.

  Thus, the predetermined amount by which the gear ratio of the toroidal type continuously variable transmission 4 is corrected based on the displacement of the spool 63 of the second actuator 61 is, for example, a GN value (based on a brake or the like on the output shaft 9 of the continuously variable transmission). Even when a large load is not applied, the driving force (creep force) is such that the vehicle can be started in the traveling direction and run at a low speed from a value that can realize a state in which the output shaft 9 is stopped while the input shaft 3 is rotated. Is an amount corresponding to an amount to be changed to a value output from the output shaft 9. Specifically, when the shift lever is operated in the P, N, D, and L ranges and the spool 63 is displaced in the other axial direction, the output shaft 9 is started in the forward direction and traveled at a low speed. The gear ratio is corrected so that the driving force can be output. Similarly, when the R range is operated and the spool 63 is displaced in one axial direction, the gear ratio is corrected so that a driving force can be output from the output shaft 9 so as to start in the backward direction and travel at a low speed. To do. Then, by performing such a gear ratio correction by the hydraulic second actuator 61 having excellent responsiveness, the required driving force can be output quickly and reliably from the output shaft 9. . In addition, in order to control the speed ratio at the start of the vehicle by the second actuator 61, the electromagnetic valve 19, the correction control valves 24a and 24b, the forward / reverse switching, which are incorporated in the conventional structure shown in FIG. The valve 65 (see FIG. 5) and the like can be omitted, and the valve body incorporating such a member can be further downsized, the apparatus can be simplified, and the cost can be reduced. The structure and mechanism for correcting the gear ratio at the start of such a vehicle are described in detail in Japanese Patent Application No. 2005-143583, and are not related to the gist of the present invention. Is omitted.

  FIG. 3 shows a second embodiment of the present invention corresponding to claims 1 to 4. In the case of the first embodiment described above, the hydraulic pressure proportional to the absolute value of the hydraulic pressure difference (differential pressure) existing between the pair of hydraulic chambers 36 a, 36 b constituting the actuator 13 is directly taken out through the oil passage 39. In contrast to the introduction into the first pilot chamber 32a of the pressing force adjustment valve 29a, in the present embodiment, a hydraulic sensor 66 is provided in the take-out oil passage 39a. Based on the differential pressure detected by the hydraulic sensor 66 (hydraulic pressure proportional to the absolute value thereof), the valve opening pressure of the pressing force adjusting valve 29b is adjustable. That is, the detection signal of the hydraulic sensor 66 is input to the controller 16, and the controller 16 controls the differential pressure (hydraulic pressure proportional to the absolute value thereof), the gear ratio of the toroidal continuously variable transmission 4, the toroidal The optimum pressing force to be generated by the pressing device 14 based on the temperature of the lubricating oil (traction oil) existing inside the continuously variable transmission 4 and the rotational speed of the engine 1 (see FIG. 4) as a driving source. Is calculated.

  Then, in order to generate the optimal pressing force calculated in this way by the pressing device 14, the open / close state of the line pressure control electromagnetic switching valve 18 for adjusting the valve opening pressure of the pressing force adjusting valve 29b, Control (duty ratio control) is performed by the controller 16. The pressing force generated by the pressing device 14 is adjusted by appropriately adjusting the hydraulic pressure introduced into the first pilot portion 32b of the pressing force adjusting valve 29b based on the opening / closing of the electromagnetic pressure control valve 18 for controlling the line pressure. Are appropriately regulated according to the operation status of the toroidal-type continuously variable transmission 4. In the present embodiment, the line pressure control electromagnetic on-off valve 18 is a normally closed type, but may be a normally open type. However, when the normally closed line pressure control electromagnetic on-off valve 18 is used, even if electricity is not sent for some reason (failed), the first pilot portion 32b of the pressing force adjusting valve 29b is used. Can be communicated with the oil reservoir 28. As a result, the valve opening pressure of the pressing force adjusting valve 29b can be secured (although it is maximum), and the minimum travel can be continued.

In the case of the present embodiment configured as described above, one hydraulic pressure sensor 66 is provided in the take-out oil passage 39a, and the differential pressure is directly detected by the single hydraulic pressure sensor 66. For this reason, there is an error in the calculated value as in the case where the hydraulic pressure in each of the hydraulic chambers 36a and 36b is detected by a pair of hydraulic pressure sensors and the differential pressure is calculated from the hydraulic pressure in each of the hydraulic chambers 36a and 36b. Can be prevented. Further, it is not necessary to filter the detection signal of the hydraulic sensor 66, and the differential pressure can be obtained quickly and accurately. Further, since the differential pressure is not directly introduced into the pressing force adjusting valve 29b through the take-out oil passage 39a, the oil passage can be further omitted as compared with the first embodiment, and the pressing force adjusting valve 29b can be omitted. The pilot portion and the spool can be reduced, and the pressing force adjusting valve 29b can be reduced in size and simplified. As a result, the valve body can be reduced in size and cost.
Other configurations and operations are the same as those of the first embodiment described above, and thus redundant description is omitted.

  In addition, although the present Example demonstrated the case where the pressing force (opening pressure of the pressing force adjustment valve 29b) of the press apparatus 14 was adjusted based on the differential pressure (hydraulic pressure proportional to) which the hydraulic pressure sensor 66 detected, The transmission ratio of the toroidal continuously variable transmission 4 can be adjusted based on the differential pressure detected by the hydraulic sensor 66 (the hydraulic pressure proportional to the differential pressure). Even in such a case, the gear ratio can be adjusted on the basis of the differential pressure (proportional to the pressure) detected quickly and accurately. For example, gear ratio control (so-called GN control) performed at the start of the vehicle, etc. Can be done properly.

  In the case of the above-described first embodiment, the low-speed and high-speed clutches 7 and 8 are connected to and disconnected from the high-speed clutch and low-speed clutch by the shift solenoid valve 20 via the shift switching valve 67. This is done by switching the switching state of the valves 25 and 26 (see FIG. 1). On the other hand, in the case of the present embodiment, the low speed and high speed clutches 7 and 8 are connected and disconnected by the low speed clutch and high speed clutch electromagnetic switching valves 68 and 69, respectively. This is done by switching the switching state of each switching valve 25, 26. The structure, mechanism, etc. for switching the connection / disconnection state of each of the low speed and high speed clutches 7 and 8 are described in detail in the aforementioned Japanese Patent Application No. 2005-143583. Detailed description will be omitted.

  In the above description, the present invention is combined with a toroidal continuously variable transmission and a planetary gear type transmission, and the rotation state of the output shaft is corrected with the input shaft rotated in one direction. The case where the present invention is applied to a continuously variable transmission equipped with a mode (low speed mode) capable of realizing a so-called geared neutral state that can be switched between rotation and reverse rotation has been described. However, the present invention combines a toroidal type continuously variable transmission and a planetary gear type transmission, a mode in which power is transmitted only by the toroidal type continuously variable transmission (low speed mode), and a planetary gear type which is a differential unit. The present invention can also be applied to a continuously variable transmission having a so-called power split state (high speed mode) in which main power is transmitted by a transmission and the gear ratio is adjusted by the toroidal continuously variable transmission. . Further, it can be used not only as an automatic transmission for automobiles but also as a transmission for various industries. The structure of the toroidal continuously variable transmission may be either a half toroidal type or a full toroidal type. Furthermore, the present invention can also be applied to a continuously variable transmission configured by a toroidal type continuously variable transmission alone.

The hydraulic circuit diagram similar to FIG. 5 which shows Example 1 of this invention. The A section enlarged view of FIG. The hydraulic circuit diagram similar to FIG. 1 which shows Example 2 of this invention. The block diagram of the conventional continuously variable transmission. The hydraulic circuit diagram incorporated in this continuously variable transmission.

Explanation of symbols

DESCRIPTION OF SYMBOLS 1 Engine 2 Damper 3 Input shaft 4 Toroidal type continuously variable transmission 5 Planetary gear type transmission 6 Clutch device 7 Low speed clutch 8 High speed clutch 9 Output shaft 10 Input side disk 11 Output side disk 12 Power roller 13 Actuator 14 Press device DESCRIPTION OF SYMBOLS 15 Gear ratio control unit 16 Controller 17 Stepping motor 18 Line pressure control electromagnetic on-off valve 19 Solenoid valve 20 Shifting solenoid valve 21 Control valve device 22 Gear ratio control valve 23 Differential pressure cylinders 24a, 24b Correction control valve 25 High speed clutch Switching valve 26 Low speed clutch switching valve 27, 27a, 27b Oil pump 28 Oil reservoir 29, 29a, 29b Pressure adjusting valve 30 Low pressure adjusting valve 31 Manual hydraulic switching valve 32, 32a, 32b First pilot section 33, 33a Second pilot section 34 Third pie Lot part 35 Pistons 36a, 36b Hydraulic chamber 37 Differential pressure take-out valve 38a, 38b Oil passage 39, 39a Take-out oil path 40 Direction switching valve 41 Cylinder hole 42 Spool 43 Source oil pressure introduction port 44 Pressure introduction path 45 First differential pressure feed Port 46 Second differential pressure delivery port 47a, 47b Drain port 48 First hydraulic pressure introduction port 49 Second hydraulic pressure introduction port 50a, 50b Second pressure introduction path 51 First reaction force port 52 Second reaction force port 53 First reaction force Chamber 54 First hydraulic pressure introduction path 55 Second reaction force chamber 56 Second hydraulic pressure introduction path 57 Return spring 58 Check ball 59 Cylinder hole 60 Sleeve 61 Second actuator 62 Hydraulic chamber 63 Spool 64 Spring 65 Forward / reverse switching valve 66 Hydraulic sensor 67 Shift switching valve 68 Electromagnetic switching valve for low speed clutch 9 electromagnetic switching valve for high-speed clutch

Claims (4)

  A plurality of supports that are supported so as to be able to swing and displace around a pivot that is twisted with respect to the central axes of the first and second disks that are concentrically arranged and relatively rotatable. A member, a plurality of power rollers sandwiched between the inner surfaces of the first and second disks facing each other in a state of being rotatably supported by each of these support members, and a piston A hydraulic actuator for displacing the piston in the axial direction based on supply / discharge of the pressure oil to and from the first and second hydraulic chambers, and displacing the support members in the axial direction of the pivot, and any one of the actuators A toroidal-type continuously variable transmission having a differential pressure extracting valve for extracting hydraulic pressure proportional to the difference between the hydraulic pressures existing between the first and second hydraulic chambers constituting the 1 extraction oil passage Through it, the first, toroidal type continuously variable transmission is characterized in that retrieving the hydraulic pressure in proportion to the absolute value of the difference between the hydraulic pressure existing between the adjacent second double hydraulic chamber.   The differential pressure take-off valve is a spool that is displaced according to the difference in hydraulic pressure between the first and second hydraulic chambers, a source hydraulic pressure introduction port for introducing hydraulic pressure generated by the hydraulic power source, and the first When the hydraulic pressure in one hydraulic chamber is higher than the hydraulic pressure in the second hydraulic chamber, the first differential pressure feed port communicating with the source hydraulic pressure introduction port and the hydraulic pressure in the second hydraulic chamber are also in the first hydraulic chamber. A second differential pressure delivery port communicating with the source oil pressure introduction port when the hydraulic pressure is higher than the hydraulic pressure, and the difference between the hydraulic pressures at that time among the first and second differential pressure delivery ports. The toroidal continuously variable transmission according to claim 1, wherein a port communicating with the source oil pressure introduction port is connected to the take-out oil passage in a freely communicable manner.   The toroidal continuously variable transmission according to claim 2, wherein the first and second differential pressure feed ports and the take-out oil passage are connected via a ball-type direction switching valve. Depending on the oil pressure of the take-out oil path detected by the oil pressure sensor, the transmission ratio of the first and second disks and the pressing device that presses the first and second disks in a direction approaching each other. The toroidal continuously variable transmission according to any one of claims 1 to 3, wherein at least one of the pressure and the pressure is adjustable.

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